• Author / Uploaded
• Abhishek Gaurav

Citation preview

I

I

1

r

i

a Group of Authorities

I

I

I

i

I i i

ROY L. HARRl NGTON

I

I

I

I

Engineering Technical Pepartmelit Newport News Shipbuilding and

I

Published by

THE S ~ C I E OF ~ Y NAVAL ARCHITECTS AND MARINE ENGINEERS

I I I

One World Trade Center, Suite 1369, New York, N.Y. 10048

i

.

@ Copyright 1971 by The Society of Naval Architects and Marine Engineers Library of Congress Catalog Card No. 78472362 Printed in the United States of America Second Printing 1976 TMrd Printing 1980

Since 1942 and 1944 when the two volumes of MARINEENGINEERING were published, the basic body of knowledge constituting marine engineering has greatly increased. Recognizing was substantially out of date, the Society in 1964 underthat the original MARINEENGINEERING took the task of compiling a reviged edition. That same year a Control Committee was ap-' pointed by the president to guide the revision, carrying on the objective of the original work, that of producing a comprehensive treatise reflecting the important technical progress of the last several decades. Also, the intent is that this text should complement the Society's two companion volumes, Principles of Naval Architecture and Ship Design and Construction, which deal similarly with the subjects of naval architecture and ship construction practices. When the task of revising the original MARINEENGINEERING was undertaken, it was quickly found to be considerably larger in scope than anticipated. The original text had to be completely rewritten, not simply revised. At the putset, it was decided that, for ready use and reference, the text should be a single volume limited to about 850 pages. Therefore, discussion of engineering subjects.covered in other textbooks had to be greatly abbreviated. Every effort has been made, however, throughout the text to make reference to appropriate source material for the individual or self-taught reader as well as the resourceful teacher (who may in some cases prefer to use his own references). Each chapter is written by a separate author (or authors). The committee felt that this precept should be continued because of the advantages of professional specialization it affords. Some unevenness in style results, but this has been minimized by the technical editor. In May, 1968, Mr. Roy L. Harrington was selected as technical editor by the committee. Mr. Hanington received a Society scholarship in 1960 to pursue an M.S. degree in marine engineering and also has had twelve years of technical ship design experience in a major shipyard. With this background, plus his extensive literary capability, he was considered well equipped to bridge the academic and the practicing professional points of view of the Society members. This book is not intended to be either a handbook or conversely a definitive text on any specific engineering discipline which may be used in marine engineering. Its purpose is to acquaint a person already familiar with basic engineering fundamentals with the various engineering disciplines and applications which constitute marine engineering. The need for such a book becomes apparent when it is recognized that many practicing marine engineers have had little formal education in the field of marine engineering as such, but instead have come into it from other related engineering activities. The Control Committee appointed to guide the revision of MARINEENGINEERING consisted of: Ernst G. Frankel Jens T. Holm William E. Jacobsen John R. Kane John H. Lsscaster ' Lauren S. McCready Andrew I. McKee Laskar Wechsler John B. Woodward I11 Robert E. Yohe

There have been so many technological advancements since the original MARINEENQINEERwas published that the'content of this book bears little similarity to the original text. For example, in a manner of speaking, a nuclear power chapter has been substituted f o ~the old reciprocating steam engine chapter, and other differences are almost as dramatic. However, the same basic philosophy was used in writing both works except that, insofar as practicable, this text covers naval practice in addition to merchant practice. In order to ensure that this book is comprehensive and factual, and accurately represents the consensus of opinion of the marine industry as a whole, the chapters were subjected to a series of reviews. After the manuscripts were prepared by the authors and reviewed within their respective organizations, they were then reviewed by the editor, Control Committee, and selected members of the marine industry who were experts in each particular area. The entire Sociehy owes a large debt of gratitude to this last group as they were largely responsible for transforming good manuscripts into excellent manuscripts. With few exceptions, it is a gross injustice to suggest that the chapters have been prepared by only the authors indicated. In several cases, the contributions of single individuals who assisted were almost as large as that of the author; and in $1 cases, the comments and discussion provided by the Control Committee and other members of the marine industry were an invaluable asset. Mr. John Markert (author of Chapter 19) accurately expressed the sentiment of the chapter authors when he stated that the generous cooperhtion and assistance received from the numerous contributors, often persons not acquainted with the author, were a revelation; it should, however, be noted that such cooperation is characteristic of the marine fraternity. An accurate listing of those who assisted in the preparation of this book would include many names. Several hundred people made direct contributions (by assisting in the preparation of manuscripts, supplying reference material, reviewing manuscripts, or supplying illustration material); and when those who made indirect contributions are added, the number of names would become even larger. - It is, however, considered proper to acknowledge some of the contributions as follows: Mr. Catlin (Chapter 3) acknowledges the valuable contributions of Mr. George W. Kessler, vice resident. Babcock & Wilcox. Mr. L. E. Triggs, chief engineer, Marine Dept., Combustion ~ n ~ i h e e r i~nc., n ~ , Mr. W. I. signell, chief marine engineer, J. J. Henry Co., Inc., and Professor J. T. Holm, Webb Institute of Naval Architecture, in the development of the chapter dealing with boilers and combustion. Dr. Illies (Chapter 8) states that he received help from a large number of individuals while preparing the low-speed directrcoupled diesel engine chapter. The material that was made available by diesel engine manufacturers (MAN, Fiat, Sulzer, Burmeister and Wain, Gotaverken, Stork, and Doxford) was particularly helpful as was the valuable advice and personal assistance that ww provided by Mr. Klaus Knaack. Mr. Semar (Chapter 9) acknowledges the contributions made by Mr. W. S. Richardson, the Falk Corporation, Mr. Norman A. Smith, General Electric Company, and Mr. Frederic A. Thoma, DeLaval Turbine, Inc., in the development of the chapter on reduction gears. Mr. J. F. Sebald (Chapter 13) acknowledges the valuable contributions made by Mr. P. D. Gold of the Worthington Corporation, Mr. William J. Bow of the Foster Wheeler Corporation and Mr. J. J. Biese of the Ingersoll Rand Co. h providing illustrations and for their critical review of the manuscript. The cooperation of the Heat Exchange Institute and The American Society of Mechanical Engineers in permitting the publication of technical data and the technical support provided by Gilbert Associates, Inc. are also gratefully acknowledge4. Messrs. Smith and Nickerson (Chapter 16) gratefully acknowledge the assistance provided by Mr. A. Taplin of the Naval Ship Engineering Center, who prepared the active fin stabilizer section of the hull machinery chapter. Mr. Stephenson (Chapter 18) gratefully notes that the machinery arrangement illustrations and many of the piping diagram illustrations in the piping systems chapter were included with the permission of Mr. W. L. Baptie of American Mail Line, Ltd. The typical chapter author is a highly competent engineer who enjoys his field of specialization and has devoted the majority of his life to it. By studying the various chapters, it will become apparent that a book such as this is published only once per generation. INQ

\

,

I -

, $

t I

1

Division 1 Introductory

1'

Chapter I

J. R.

GENERAL CONSIDERATIONS I N MARINE ENGINEERING

KANE, Director of Engineering,

Newport News Shipbuilding and Dry Dock Company

1. Intrbduction .......................... 2. Concepts and Concept Formulation.. .... 3. Ship System Formulation.. ............. 4. Development of Main Propulsion System Requirements.. .....................

%. ~ a i Propulsion n Plent ~rade-offStudies. 6. Preliminary Design Considerations. . . . . . 7. Specifications. ........................ 8. Final Design and Working Plans. . . . . . . . 8 9. Tests and Trials. ..................... 1 2 5

$

11 18 31 33 35

Power Plants

I

Chapter I1

THERMODYNAMICS AND B E A T ENGINEERING

Jws T. HOLM, Professor, Webb Institute of Naval Architecture J. B. WOODWARD 111, Professor, University of Michigan

I

1. Review of Fundamentals. .............. 2. Heat Transfer in Boilers. . . . . . . . . . . . . . . 3. Internal Thermodynamics of the Steam Tutbine. ...........................

Page PW~ 38 4. External Thermodynamics of the Steam Turbine ............................ 55 5. ~herniod~namics of steam Cycles. . . . . . . 61 / 49 6. Waste Heat from Diesel and Gas Turbine i Engines ............................ 73

/

i b

Chapter I11

BOILERS AND COMBUSTION

EVERETT A. CATLIN,~ a x i n Engineer, e The Babcock & Wilcox Company 1. Classification of Marine Seam Generatom 2. Considerations in the Selection of a Boiler

Chapter I V

pa@ 78 3. Boiler Pesign . . . . . . . . . . . . . . . . . . . . . . . . . 94 90 4. Boiler Operation.. ..................... 125

NUCLEAR MARINE PROPULSION

ROBERT T. PENNINQMN, formerly Manager of Nuclear Maxine Engineering, Advanced Products Operation, General Electric Company

I

Page page 1. Basic Fundamentals. .................. 130 3. . Nuclear Propulsion Applications. . . . . . . . 149 2, Reactor Design Considerations. . . . . . . . . 138 vii

I STEAM TURB1,NES WILLIAMI. H. BUDD,Assistant to Manager of Engineering, Marine Systems, DeLaval Turbine, Inc.

unrtpucr v

Turbine Control.. .................... Rotors and Blades. .................. Norzlea, Diaphragms, and Stationary Blading.. .......................... Casings &adPackings. ................ Lubrication and Bearings. ............ Main Propulsion Turbine Operation. . . . Auxiliary Turbines. ..................

1. Nonreheat Main Propulsion Turbines. .. 2. Reheat Main Propulsion Turbines. ..... 3. Main Propulsion Turbine-Nuclear

Cycle. ............................. 4. Combined Steam and GaB Turbine Main Propulsion Cycles. ................. 5. Turbine Speed, Number of Stages, Dimensions. ....................... : .

1

Chapter X

ELECTRIC PROPULSION DRIVES

W. E. JACOBBEN, Manager, Marine Systems Engineering, General Electric Company 1. Introduction.. ........................ 334 2. The Diesel Direct-Current Drive System. 339 3. The Turbine Direct-Current Drive System 347

180 185

4. The Turbine Alternating-Current Drive

System . . . . . . . . . . . . . . . . . . . . . . . . . . . . 348

5. The Diesel Alternating-Current Drive

System. ........................... 356 6. Electric Couplings.. ................... 360

190 193 196 199

Chapter XI

201

PROPELLERS, SHAFTING, AND SHAFTrNG SYSTEM VIBRATION ANALYSIS

Assistant Chief Engineer, Newport News Shipbuilding and Dry Dock Company C. L. LONG, Chapter VI GAS TURBINES A. 0. WHITE, Manager, Advanced Applications Unit, Medium Electric Company

.

1. 2. 3. 4. 5.

Basic Considerations.. .................. Arrangement and Structural Details. . . . . . Accessories. ........................... Controls. ............................. Centrifugal Compressor Design. .........

206 213 218 219 222

6. 7. 8. 9.

Page

~k Turbine

Operation, General

Axial-Flow Compreseor Design. . . . . . . . . Turbine Design and Construction. ...... Combustion Systems. ................. Bearings, Seals, and Lubrication. .......

Introduction.. ........................ 362 b5. 2. Arrangement Considerations.. . . . . . . . . . . 365 e 6 . p - 3 . Shafting Loads.. . . . . . . . . . . . . . . . . . . . . . . 366 -7. ' '/4. Shafting Design. . . . . . . . . . . . . . . . . . . . . . . 372 8. b- 1.

I

1. Introdrtction. ......................... 246 2. aaracte$tics of Diesel Engines. . . . . . . . 251

Bearings.. ............................ Propellers. ........................... Torsional Vibration.. .................. Longitudinal Vibration.. ...............

379 384 388 393

9. Whirling Vibration.. ................... 397

225 229 235 239

Division 4 Auxiliary Co~aponents

MEDIUM AND HIGH-SPEED DIESEL ENGINES LASKARWECHBLER,Technical Director, Machinery Systems ~ i v i s i o n ,Naval ship ~ n ~ i n e e r i n ~

v

Page

PUMPS, FORCED-DRAFT BLOWERS, COMPRE$SORS, AND EJECTORS Supervisor, Centrifugal Pump Engineering Departmen;t, DeLaval Turbine, Inc. G. W. SOETE, Page

page

L=-

3. Marine Uses for Diesel Engines. . . . . . . . . 257 4. Design Considerations. ................ 261

1. Centrifugal Pumps.. ................... 401 4. Rotary Pumps.. ...................... 432 2. Reciprocating Steam Pumps. ........... 422 \.-5. Forced-@aft Blowers.. . . . . . . . . . . . . . . . . 436 3. Power Pumps.. ....................... 428 L. 6, Compressors.. ......................... 440

7. Ejectors.. ............................ 444

LOWSPEED DIRECT-COUPLED DIESEL ENGINES KURTILLIES, Professor, Technische Universitat Hannover Pege 1. survey of Principal &acteri&ics. ..... 280 2. Engine Subsystems. ................... 292

Chapter XI11

JOBEPH F. SEBALD, Consulting Engineer and Special Consultant to Gilbert Associates, Inc.

3. Overall Considerations. ................ 303

Transmissions REDUCTION GEARS

HAROLD W. SEE~AR, Manager, Technical Support, Marine Mechanical Dep&ment, westinghouse

MAIN AND AUXILIARY CONDENSERS

!*

Page

Page

1. General C~aracterhtics................. 450 2. Condenser Design. . . . . . . . . . . . . . . . . . . . . 456

3. Surface Condenser Performanm. . . . . . . . . 473 4. Performance Predictions from Design Geometry.. . . . . . . . . . . . . . . . . . . . . . . . . 478

Chapter X I V

HEAT EXCHANGERS

CHARLEB D. ROBE,Vice President, AquaXhem, Incorporated PHILIPLIU, Chief Thermal Design Consultant, Research and Development, Aqu*Chem, corporated

Electric Corporation page v

1. Introduction.. . . . . . . . . . . . . . . . . . . . . . . . . 310 3. Gear Design. . . . . . . . . . . . . . . . . . . . . . . . . . 317 2. Tooth Design Factors. . . . . . . . . . . . . . . . . 313 v 4 . Applications.. . . . . . . . . . . . . . . . . . . . . . . . . 331

page

page

1. . Introduction. ......................... 488 2. Heat Transfer in Shell-and-Tube Heat Exchangers.. 496

3. Heat Exchanger Applications. .......... 514

In-

.......................

ix

Chapter XV

DISTILLING P U N T S

.

Division 6 Supporting Technology

C ~ I D.WROBE,Vice President, AqueChem, Incorporqted page

Page

1. Distilling Plant Designs. ............... 530

2. Distilling Plant Design Considerations. .. 550

Chapter XVI

-

HULL MACHINERY

IRVING W. SMITH,Mechanical Engineer, Office of Ship Construction, Maritime Administration ARCHERM. NICKERBON, JR.,Senior Engineer, J. E. Bowker Associates, Inc. v

Chapter XX

1. General Design Consideratioqs. .........

ptlge 564

L-

2.

'

570

WATT V. SMITH, Head, Friction and Wear Branch, Materials Department, Naval Ship Research and Development Laboratory, Annapolis, Maryland J. M. GRUBER,Vice President, Waukesha Industries Corporation

I I

,

1

page

Hull Machinery Installations.. ..........

page

I

ELECTRIC PLANTS

Chapter XVIII

1. Machinery Space Arrangement.. ........ 670

'2. Piping Design Details.. ................ 676 ;'3. Piping Systema........................ 682 "

ENVIRONMENTAL CONTROL

JOHN W. MARKERT, Professional S u p p o r t A i r Conditioning, Office of Construction Management, Public Buildings Service, General Services Administration 1. 2. 3. 4.

Page 2. Applications.. ........................ 796

CONSTRUCTION MATERIALS

W. LEE WILLIAMB, Assistant Bead, Materiala Department, Naval Ship Research and Development Laboratory, Annapolis, Maryland M. ROBERTGROSS,Head, Materials Engineering Branch, Materials Department, Naval Ship Research and Development Laboratory, Annapolis, Maryland 1. 2. 3. 4.

Prefacing Remarks. . . . . . . . . . . . . . . . . . . . Corrosion of Metals.. .................. Fatigue ............................... Behavior at Elevated Temperatures. ....

Page

Page

810 810 817 821

5. Applications of Materiala.. .............. 824 6. Glossary of Metallurgical Terms Used in Materials Engineering.. .............. 835

PIPING SYSTEMS

Manager, Piping Design '~epartment,Newport News Shipbuilding and Dry E. E. STEPHENBON, Dock Company ',

Pa%e 1. Automation System.. ................. 791

Chapter XXII

6. Lighting and Power Distribution. ... :... 640 7. Interior Communications.. ............. 654 8. Electronic Navigation and Radio Communication..................... 659 9. Wiring Application and Methods. ....... 663

605 607 614 621 635

AUTOMATION

W. 0. NICHOLS,Chief Engineer, Central Technical Division, Shipbuilding Department, Bethlehem Steel Corporation

Division 5 Sl~ipboard Systems

Introduction.. ........................ Generating Plants. .................... Switchboards and Panels. .............. Powe~Equipment. .................... Lighting Fixtures and Equipment. ......

Page

1. Review of Fundamentals. . . . . . . . . . . . . . . 770 ""2. Bearings.. ............................ 778 \--~ 3. Lubrication System. ................... 785

Chapter XXI

1. 2. 3. 4. 5.

BEARINGS AND LUBRICATION

Intraduction.. ........................ System General Requirements. ......... Design Criteria and L o 4 Components. .. Piping Systems.. ......................

710 718 726 734

5. 6. 7. 8.

Air Handling System Resign.. .......... Air Handling Equipment. .............. Beating and Cooling Equipment. ....... Refrigeration Equipment.. .............

745 756 763 766

Chapter XXIII

PETROLEUM FUELS

Manager, Technical Services, Marine Sales Department, Mobil Sales and CARLE. HABERMANN, Supply Corporation Page

Page

1. Fuel Manufactureand Characteristics. ... 842

2. Fuel Procurement. . . . . . . . . . . . . . . . . . . . . 853

page INDEX.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 858

CHAPTER I

J. R. Kane

1

I

General Considerations in Marine Engineering

Section 1 Introduction The first efforts to apply mechanical power to the propulsion and operation of ships date back to the early oighteenth century, nearly concurrent with the start of the Industrial Revolution. By the beginning of the nineteenth century, almost a full century before the Wright brothers made their first sporadic flights in a glider at Kitty Hawk, ~ t e a m - ~ r o ~ e ships ied had become a commercial reality, and marine engineering was born. Considering such an early beginning, it mems paradoxical to have to say now, well along in the twentieth century, that it is difficult, if not impossible, to write a definitive text on the subject. Such is the case, however, since the field continues to enter new oras of activity and evolution. One of the reasons this subject is difficult to treat is that ships have never been simple products but, to the oontrary, require an exceptional number of specializations to plan, design, and build. Thus marine engineering is not as simply categorized as, for example, civil, machanical, electrical, or chemical engineering, but is an integrated engineering effort comprising parts of many ongineering disciplines directed to the development and dosign of systems of transport, warfare, exploration, and tlstural resource retrieval which have only one thing in earnmon; namely, that they operate in or upon the crurface of a body of water. The field of engineering activity designated as naval wrohitecture and marine engineering is concerned with at let~stthe following areas:

I

Ocean engineering. The conception, design, construction, and operation of vehicles, submersibles, and fixed or floating structures and their integration into systems for the conduct of oceanographic research, exploration of ocean resources, and the utilization of ocean resources are encom~assedin this categorv.

The division of responsibilities between the naval architect and the marine engineer differs from one activity to another. However, the marine engineer is, in general, responsible for the engineering systems required to propel, work, or fight the ship. More specifically, the marine engineer may be responsible for the main propulsion plant, the powering and mechanization aspects of ship functions such as steering, anchoring, cargo handling, heating, ventilation, air conditioning, electrical power generation and distribution, interior and exterior communication, and other related requirements. The naval architect, in general, is primarily concerned with the hydrodynamic and hull form characteristics of the ship, the structural design of the hull, the control aspects of the vehicle, habitability considerations and the ability to survive and endure in the service environment. The naval architect, assisted in appropriate areas by the marine engineer, is responsible for the overall arrangement or configuration of the ship extending to both the exterior and interior arrangements. I n addition, the naval architect is generally charged with the responsibility for the overall esthetics of the design, the interior decoration, and the general suitability and pleasing Inland waterway and ocean transportation. The con- quality of the architecture. usplion, design, construction, and operation of vehicles Certain aspects of the design of marine vehicles are utilizing the waterways and oceans, especially the ocean difficult gto clearly assign as the responsibility of either murfaces, for transportation of commodities, goods, and the naval architect or the marine engineer. The design personnel, are included in this category. The integration of propellers or propulsors is one of these, being in the of tho operation of these vehicles with land transport& minds of some a hydrodynamic device in the domain of tion via harbor and terminal facilities is an extremely the naval architect, and in the minds of others to be an hportant consideration. In the case of small boats, energy conversion device similar to pumps, turboymbts, and cruise ships, transportation may be secondary machinery, and the like, thus in the sphere of the marine h lsi~ureor sport as an objective. engineer. Hull vibration, excited by the propeller or by Naval engineering. This category includes the con- the main propulsion plant, is another such area. Noise aegt,ion, design, construction, and operation of naval reduction and shock hardening, in fact dynamic response rtcrfaoo ships and submarines and their integration into of structures or machinery in general, usually must be wsrf~bre systems. Means of appraising the military the joint responsibility of both the naval architect and effrotivenessof these systems and the optimal utilization the marine engineer. Cargo handling, cargo pumping of thoir properties are major considerations. systems, environmental control, habitability ,. hotel

2

I

MARINE ENGINEERING

1

services, and numerous other such aspects of ship design all involve joint responsibility and interfacing between the naval architect and the marine engineer. The traditional distinctiowbetween naval architecture and marine engineering in t k multifarious aspects of ship selection, design, construction, and operation are tending to disappear, to be replaced by broader concepts of systems engineering and analysis. Because of the multidisciplined nature of marine engineering and naval architecture, they have been particularly affected by the impact of the explosive growth of technology during recent years. Prevalent use of the electronic computer has been particularly influential, in that complex rnathematical analyses once considered prohibitively laborious are now routinely made. By providing the ability to rapidly conduct an increased number of computations, readily store and analyze data, and simultaneously

consider a larger number of factors, the computer makes mathematical simulation of complex problems feasible and is leading to a better optimization of designs. Furthermore, due to the period of large-scale industrial development into which we have entered, there is increasing acceptance of the principle of planned technology which affects systems of all sorts, including marine transportation, oceanography, and recovery of ocean or ocean-bottom resources. By surveying the series of inventions and innovations which have established the present state of the art of marine engineering, it becomes apparent that engineering in the ocean environment is characteristically a dynamic, continuously advancing technology. As a result, this text must be considered an interim report of the processeis that are developing in a broadening marine engineering field.

I

I

I

Section 2 Concepts and Concept Formulation 2.1 Early History. In about the year 1712, an enterprising blacksmith from Dartmoor, England, by the name of Thomas Newcomen, successfully developed a rudimentary steam engine for the purpose of pumping water out of mines. This engine consisted essentially of a single-acting piston working in a vertical open-topped cylinder. The piston was packed with hemp since the state of the metal-working art was very primitive and a tolerance of about one-sixteenth inch out of round or "the thickness of a thin sixpence" was about the best that could be expected. The piston was connected to one end of a rocker arm by a chain without a piston rod or guide. The differential working pressure was derived primarily from the vacuum which was created below the piston by water spray into the steam space a t the end of the upstroke. The steam and water valves were worked by hand. Some sixty years later, radical improvements were made by James Watt, whose name is more frequently associated with the early development of the steam engine. I n the course of time, numerous other.improvements followed, of which the most important was probably the double-acting inverted vertical engine which proved to have so many advantages that it has remained standard ever since. Accounts of the work of men such as Savery, Newcomen, Papin, and Watt in connection with the invention and development of steam engines are truly exciting [I, 2,3].l Despite the much earlier development of steam engines, their application to the propulsion of ships was not undertaken until about 1784. Attempts to adapt the early steam engines to ship propulsion were carried out almost simultaneously in America, Scotland, Numbers in brackets designate References at end of chapter.

and France, and a t least seven reasonably practical steamboats were developed before 1807 when Robert Fulton inaugurated the first commercially successful use of steam marine propulsion in the small wooden paddle wheel vessel Clermont [I]. The Clermont operated up the Hudson River from New York to Albany, a distance of 150 miles, in about 32 hr. Although paddle wheel vessels were promptly adopted for river service, twelve years elapsed after the launching of the Clermont before the steamer Savannah made the first ocean voyage from America to Europe. It should be noted, however, that even in this instance the machinery was not operated continuously during the outbound leg of the trip and the inbound leg was made under sail. The era of the paddle wheel steamships reached a climax about 50 years later when the steamship Great Eastern was built. This was a steel-hulled vessel almost 700 f t long and 22,000 tons burden, which is large even today for a cargo vessel, and which had the principal fault that it was too advanced for its time. The introduction of the screw propeller in 1837, which was a revolutionary development, similarly did not immediately displace sailing vessels. As late as 1860 the speed of the best clippers still exceeded that of any steams hi^ and the greater d art of the work a t sea continued td be accomilished inder sail. B y the year 1893, the year of the founding of The Society of Naval Architects and Marine Engineers, the screw propeller.. driven by a triple-expansion steam engine had become the,predominant means of propulsion of seagoing ships although t addle wheels were still used with river-and- excursion steamers. Steam was almost universally produced by Scotch boilers and coal was the

GENERAL' CONSIDERATIONS

c o w o n fuel. The steam turbine and diesel engine were yet to make their debut. The decade from 1893 to 1903 was a period rich in marine engineering development. The early reciprooating steam engine reached the point of development of the six-cylinder quadruple-expansion engines of 10,000 indicated horsepower supplied with steam by Scotch boilers a t 200 pounds pressure. The use of electric power generated by engine-driven "dynamos" a t 100 to 112 volts was increasing rapidly. Water tube boilers, which would eventually replace the Scotch boiler on the seas, had become established in England and in the United States. An important milestone in marine engineering was the development, by Sir Charles A. Parsons, of the first successful application of the steam turbine for marine propulsion; this was accomplished aboard the Turbinia, a small vessel similar to a torpedo boat. The rotative speed of the Turbinia's three series turbines was about 2000 rpm, and they were coupled directly to relatively primitive screw propellers in a triple shaft arrangement. Parsons was dismayed on his earliest trials to discover that the wheels more or less "bored a hole in the water," developing disappointingly low driving thrust. Much developmental work was necessary before this new prime mover was successfully adapted to the requirements of marine propulsion. In what must certainly be considered one of the earliest efforts at model tank testing of propellers, Parsons investigated the subject of cavitation and succeeded in redesigning his propellers (three per shaft were ultimately employed) such that in 1897 a t a naval review of the British fleet a t Spithead, England, the Turbinia astounded the British admirals by steaming past smoothly a t a speed of 34 knots, belching smoke like an angry bull tossing dust. Lord Kelvin described this development as "the greatest advance made in steam ongine practice since the time of James Watt" [4]. Prior to 1893, a number of internal-combustion engines were attempted using anything from gunpowder to gas. One of these was a radically different type of engine in which the combustion air charge was compressed to a pressure and temperature above the ignition point of the fuel; it was patented by Dr. Rudolf Diesel, a German engineer, in 1892. There were very serious Wculties to be overcome with the diesel engine, development proceeded slowly, and it was not until fifteen to sixteen years later that a successful commercial diesel enginc of 25 hp was produced. Once this had been achieved, however, rapid progress waq made, and in a few years many firms in Continental Europe were actively building diesel engines with as much as 500 hp per cylinder. Already a t that early date experimental cylinders of 2000 horsepower were under test. The challenge to the coal-fired low-pressure reciproaating steam engine came from the steam turbine and the 'dio~elengine about the same time a t the turn of the aantury. World War I retarded developments, however, etld maintained the supremacy of coal for a little while

3

longer. After the war, oil found preference either as diesel engine fuel or for raising steam. It also reduced crew requirements and made fuel storage an easier task. The historical developments noted in the foregoing were beginnings which, when viewed against the techniques and materials available a t the time, were magnificent conceptions. No effort has been made here to include the full roster of great names and pioneer events in marine engineering. However, some familiarity with the background of the early days in marine engineering is highly recommended for those entering this field to develop an appreciation of the hopes and disappointments, the dreams and disillusionments, and the blood and sweat which lie behind the present state of the art [I-81. 2.2 Broader Concepts-Systems Analysis. The concept which motivated the majority of the early attempts in marine engineering was quite simple; namely, to develop a superior system to overcome the vagaries of the wind and the feebleness of muscle power in the propulsion of ships. However, marine engineering today entails much broader system requirements and concepts than most developments of that time. By way of introduction, one particular historical undertaking is given special note since it contained, in a primitive way, elements of systems analysis. In 1776, a year which should strike a familiar note with most Americans, a Connecticut Yaxikee named David Bushnell built the Turtle, the first submersible craft to make an undersea attack during warfare. The Turtle of the American Revolution, so called because it could be likened to two turtle shells clamped together, was built of barrel staves and iron, contained ballast tanks which were flooded to submerge, and was moved by primitive spiral screws. Reference [8] contains an interesting description of the Turtle and its precocious concepts. The Turtle was not by any means the first successful submersible craft, but was one of the most significant, since among other things it was one of the earliest, and perhaps boldest, attempts to develop a military system involving an evolutionary marine vehicle. The operational concept of the Turtle d i e r e d somewhat from most other inventions of that era since it related in a primitive way to an entire system. It was intended that the pilot dive the vessel under the water in order to evade lookouts on an enemy vessel, attach a time-delayed explosive mine to the ehip's bottom, and make a safe escape. The initial target of the Turtle was Admiral Howe's 64-gun flagship, HMS Eagle. The story of this initial venture is fascinating; the Turtle did not in fact succeed, kt.it came perilously close to doing so. George Washington wrote to Thomas Jefferson a t the time of the Turtle, "I then thought and still think that it was an effort of genius, but that many things were necessary to be combined to expect much from the issue against an enemy who are always upon guard" [91. Although the development of the first ironclads, the Merrimac and the Monitor, almost a century later had probably a more revolutionary effect on the evolution of

GENERAL CONSIDERA'I'IONS

MARINE ENGINEERING

5

I

STATE SYSTEM OBJECTIVES

DETERMINE CONSTRAINTS

-\:

r

DELINEATE SYSTEM REQUIREMENTS DELINEATE DESIGN REQUIREMENTS

- I

i

DEVELOP DESIGN ALTERNATIVES

~r \

t

\ ln

= &

PERFORM TRADE-OFF STUDIES

-ESTABLISH OPTIMUM DESiGN

-

ii

L

1L

DELINEATE DETAILED SPECIFICATIONS DETAILED

Fig. 1 - Functional processes in a systems analysis .

.

warships, Bushnell's submarine is of special interest because of the singularity of its operational concept and its primacy. Actually it contained all the elements of a modern problem in concept formulation for a planned technological development: a mission objective or primary task, an analysis of the objective to establish specific operational requirements, trade-offs concerning alternative methods of accomplishing the mission, constraints imposed by limitations of techniques, materials, manpower, money, and time, and last but not least, the necessity of obtaining the interest and support of the controlling authority for what must have seemed, in this case, to be a radical venture. In the early historical stages of the basic engineering process, the concepts formed and the decisions made, although frequently ingenious, were of sufficiently narrow scope that a single individual could become intimately familiar with all facets of the undertaking. The stakes were high for a successful development; rugged individualism was the rule since society had not yet embraced the role of technological development, and support by the existing governing bodies was scanty or nonexistent. Success depended to a large extent upon

intuitive perception and upon chance. Today, in this age of institutionalized knowledge and electronic computers, such factors are still important, but are being largely transcended by systematized approaches and by team activity. The ship, which once was viewed as a highly subjective entity, possessed of feminine and almost human attributes, is now looked upon more objectively as a link in a transportation system, a military platform, or as a medium in a system of transferring people, commodities, national presence or authority, and the like from one point to another. From a functional point of view, a ship is a most complex vehicle which must be self-sustaining in its element for long periods of time with a high degree of confidence. A ship is perhaps the most multipurpose vehicle having more built-in functions than any other type; and, as a part of a transportation or military system, the ship envelope contains a greater variety of components than any other vehicle in the system. A ship's mechanical, electrical, and structural systems are quite complex and are further complicated by the fact that they must be environmentally oriented. Due to the complexity of ships and their interfaces in transportation networks, the design of optimum ship systems cannot practicably be undertaken in a random manner. The design of complex systems involving ships is best accomplished by utiliing the systems analysis approach [lo-141 as schematically illustrated by Fig. 1. I n this way, the design process can be organized in logical steps so as to ensure that, when completed, every facet of the design has been given proper treatment. As indicated in Fig. 1, a systems analysis is initiated by establishing a system objective. Beyond that point the systems analysis approach is a continuously iterative process with each of the functional processes possibly having an impact on those remaining. For example, referring to Fig. 1, the initial system objective could be to transport cargo between two points at a given rate and a t the lowest possible cost. Proceeding with this objective, constraints such as time and capital limitations must be established. Since the constraints may alter the original objective (e.g., preclude transporting cargo at the desired rate or make higher rates attractive), the original objective must be reevaluated. The various aspects of the design process continue until all factors in the analysis are compatible, at which time the design is complete. In more general terms, a combination of theory and facts (including a careful statement of the constraints upon the system) is used to ~roducean abstract study or model of the actual situation. The model, in turn, is combined with a set of aims to produce a plan of action or a proposed technical approach. Working with such analyses and with checks against experience and data gives rise to a body of correlated information which feeds back to modify the designs which are acceptable, the facts which are relevant, the controls which are efficient, is and the aims which are realistic. Systems engidng

I

1

1

the term for such a process when limited to basic engineering processes. Systems analysis is the more general term for the process when social and economic factors in addition to basic engineering processes are included. Operations research is the name of the process when operability, that is, the optimum deployment or utilization of components, men, and machines, is the principal objective. Work study is another term of related connotation, although in this case the emphasis is placed on optimum utilization of man, and reduction in manning requirements, by taking a fresh look at work patterns and habits that have come to be taken for granted. The objectives and constraints upon which the policy for systems analyses is based have differing motivations for military systems and for merchant marine transport systems; but in both instances they ultimately reduce to the same base-cost effectiveness. I n the case of merchant systems, the proposed system must be cost effective as compared to other potential investments in order to command the necessary venture capital under the free enterprise system, or they have to be justified for governmental support by subsidy. Military planners are charged with the national defense, but there is in fact a limit to the amount of money available for such purposes as there are more military systems competing for funds than can be supported by the funds available. Consequently, the analysis of military budgets becomes a process of identifying systems, or combinations of systems, which have the maximum military cost off ectiveness. Cost effectiveness seems simple to comprehend, but usually is difficult to quantify [15]. In general, the

expression denotes a measure of the degree to which the achievement of the tasks or missions of a system (e.g., revenue earned or national protection provided) has been maximized relative to the costs associated with the system. Since the effective life of a ship is approximately twenty to twenty-five years, a period long enough for economic and political factors to undergo substantial change, the projection of life cycle costs associated with ships is inherently less accurate than life cycle cost estimates made in connection with vehicles such as automobiles or aircraft which have a much shorter life cycle. When conducting life cycle cost analyses with ships, which are relatively long lived, considerably more importance must be attached to the events which occur during the early stages of the ship's life. There is little question that the basic vehicle will perform satisfactorily for a 25-year life; however, there have been' many cases in which ships have been reequipped, modernized, jumboized, converted, etc., a number of times during their lives. As a result, the credibility of projections for the first five or ten years of a ship's life are considerably better and are often given more weight than more distant forecasts. However, despite the uncertainties associated with long-range forecasts, attempts to project them are being made and a new branch of systems analysis termed assurance engineering has been developed to give numerical expression to characteristics such as reliability, maintainability, logistic aupport, operability, safety, and similar factors which augment the standard design performance estimates traditionally made. Also, producibility analyses, requiring a combination of design and industrial engineering skills, are sometimes made to assure a design best adapted to economy in construction.

Section 3 Ship System Formulation 9.1 Mode of Utilization. Before proceeding with a mview of the marine engineering phase of a ship system formulation, which as indicated by Fig. 1 does not oornmence until the broader aspects of the system have boen tentatively formulated, it is useful to review some af the broad considerations. In particular, the modes in which ships can be utilized and the payload and speed oharacteristics of ships are of great importance in that they must be compatible with the overall system oonsiraints.. From the viewpoint of utillation, marine vehicles mny be classified in the following three categories:

in terms of deadweight and cubic requirements, must be very carefully analyzed as the latter will have a controlling effect on the vessel configuration. (b) As a mobile fighting base. Seaborne bases for force groups, weapons systems, missiles, aircraft, or other sJrstems of warfare either tactical or strategic and either offensive or defensive are included in this group. In this instance, the design of the ship is subordinated to the military system and weapon requirements except for certain inescapable essentials such as seaworthiness, habitability, etc. Payload in this case will generally be defined in military terms relating to militaw effective(a) As a link in a tramportation ~ s t e m . Inthis case, ness, and the speed requirement will be a function of the payload, mean effective speed between t e d n a l s , turn- expected speed of the hostile forces and the successful mound time, and the number of vessels are the ~rimary accomplishment of the n~ission. vmiables and must be considered in relation to their (c) As a* special-purpose vehicle or platform. This gffeot on the initial and daily operating costs as well as category includes many diversified craft which have little tho other facets of the transportation system. Payload, in common except that they all work or operate in an

,

6

GENERAL Cob

MARINE ENGINEERING I

Table 1

A Comparison of Constraints Imposed upon Merchant and Military Ship Systems

Tramportation market potentiak cargo and/or passengers Type of tran ort system contexnplated:?ulk, break bulk, containerl passengeFcar o combinatmliquid and buk etc. Most likely itine terminal facilities, h a r b ~ h t a t i o m , c d limitatiom, and fueling ports Linking services: shore d@ribution systems, new termma1 facilities, cranes, and so on Competing services Socio/political considerations and union relations Economic projections, financial support, government subaidii&etc. Technologid development, state of the art G c t o bodies, such as ABS and U%G

Fig. 2

Specitlc power Venus speed for various vehicles

ocean or waterway environment and that much support for the systematic design of them is derived from the body of marine engineering knowledge obtained from less specialized vessels. Oceangoing tugs, salvage vessels, oceanographic research ships, submersibles, dredging vessels, yachts, ferryboats, towboats, pushers, barges, hydrofoil craft, surface effect ships, and many others are examples of such special-purpose craft. Category (c) does not lend itself to generalization beyond the fundamentals of naval architecture and marine engineering. Neither, one might conjecture, do (a) and (b). However, the constraints to be considered in determining system requirements so as to ensure a reasonably optimum design configuration do parallel between merchant and military applications to rs surprising extent as indicated by the comparison in Table 1. 3.2 Payload and Speed Considerations. I n addition to the constraints dealing with the mode of utiliiation, payload and speed considerations have a strong influence on the selection of the type of vehicle employed. Payload and speed constraints are important in that they restrict the types of vehicles which are feasible for parti~ularapplications. Figure 2, parts of w h i ~ hwere taken from references [16-201, is an informative com-

Type of war situation anticipated Tactics, strategy, mission pro-

,,

Most like1 operational locale, support8aaes, replenishment means, etc. Force pou compatibility, potential dies Enemy threat in weapons and ship types Socio/political considerations Fiscal environment and budg e t pressures ~ Technolo 'cal development, state o&he art Military specifications

parison of alternative means of transportation in that the feasiblerange of speed for the various types of vehicles becomes evident. Although payload considerations are still a factor, size restrictions are less stringent in connection with ships than with the alternative modes of transportation. An investigation of a systematic family of ships (a parametric study in which size is the principle characteristic that is varied) will demonstrate that ships are not sizelimited and can be built as large as one may wish without encountering limitations from the laws of physics. Dimensional analysis will show that geometrically similar ships of a diierent scale will float at the same proportionate draft since both the water displaced (buoyancy) and the weight of the ship tend to increase as the cube of the scale. A corollary conclusion from such systematic investigations is that displacement ships are not particularly weight-sensitive. Vehicles such as fixed-wing aircraft, hydrofoil craft, planing boats, and surface effect devices in general are weight-sensitive and size-limited as may be seen from a simple dimensional analysis. Such craft derive their support in flight from lifting surfaces of various types; when geometrically similar but larger versions of a prototype are considered, the weight of the craft, including its payload, increases approximately as the cube of the scale ratio while the area of the lifting surface increases only as the square. As a result, the unit pressure loading on the lifting surface increases directly with the scale. The increase in size of fixed-wing aircraft over the last several decades has been achieved largely by increasing the forward speed by almost an order of magnitude and by greatly refining and improving the lifting character-

istics of wings and fuselages by means of extensive research developments. As the speed in flight is increased, the basic configuration of the aircraft must be changed appropriately also, because, as compared with diplacement-type ships, vehicles in the aircraft or surface-effectsupported category tend to be size-limited and weight-sensitive. As may be evident from Fig. 2, the displacement type of vessel has very definite limitations with regard to the speed at which it can be efficiently driven. The inherent speed limitations for ships are most appropriately expressed in terms of the so-called speed-length ratio (the ship's speed in knots divided by the square root of the ship's length in feet) in conjunction with various ratios of the ship's dimensions such as the beam-draft ratio and the prismatic and block coefficients (see reference [21] for a comprehensive treatment of this subject). The most spectacular growth in the size of ships has been in tankera. During the early 19509s,the so-called supertankers were in the cargo deadweight range of 20,000 to 30,000 tons; whereas during the latter 19609s, tankers as large as 200,000 to 300,000 tons were being built with projected giants in the 1,000,000-ton range appearing feasible. The theoretical problem of optimizing a transport system would appear to be simply that of maximizing payload times mean effective speed from point to point while a t the same time minimizing initial costs and yearly operating costs. If this were the only consideration, ships would be in much greater favor as compared with aircraft than they are. Systems analyses of typical transport missions usually include another highly important factor which puts a great premium on higher speed; namely, flexibility, or the ability to be in the right place at the right time with the right payload. The great increase in the speed of communications and the resultant great increase in the rapidity of affairs in recent decades has resulted in a higher premium on speed and time in many instances whether justifiable or not. Aircraft, therefore, usually transport a substantial proportion of the people, special equipment, and lighter commodities in which cases speed is of great importance, while ships continue to carry the larger proportion of the heavy cargos and commodities and bulk cargos in both military and nonmilitary transoceanic routes. 3.3 Deflnition of Fundamental Requirements. The constraints imposed by the intended mode of utilization and requirements regarding payload and speed will Ittrgely define the fundamental requirements of the ship, and an analysis of the ship system can now be conducted for the purpose of establiahing a reasonably optimum aolution. All of the positive constraints upon configuration should be identified in the analysis, but as much freedom of selection retained as possible. Once the objective and the constraints have been clearly stated, tho analysis may often proceed to the development of a u~oful abstract model for the system. Parametric &dies, in which the prin~ipalindependent variables are

varied systematically, using the electronic computer as appropriate, are often made. The sensitivity of the system to variation of the independent variables begins to emerge and can be identified. Because of its value in decision-making, the sensitivity of system characteristics to such systematic variation of the system parameters is often specifically explored in a formalized sensitivity analysis. Exercise of such techniques should result in sufficient background to support decisions regarding a policy and a plan of action. This plan of action will generally result in decisions which will further limit the range of variables to be considered; for example, the range of the size and the numbers of ships required may be more confined, notional ship design arrangements may be selected, approximate manning requirements determined, first approximation of costs projected, and so forth. A description of some of the procedures which may be used during the preliminary design of a ship is given detailed treatment in references [22-31.1. In the case of cargo ships, the fundamental concept of the cargo transportation system must be established at this point as the design of the entire system is predicated on this decision. General cargo transportation systems which employ intermodal containers (i.e., systems in which cargo is packed in containers that are transported by trucks, ships, barges, and trains in any combination before being delivered and unpacked) are becoming increasingly popular. The use of intermodal containers offers several advantages, the major one being the minimization of the number of times the cargo must be handled on an individual basis with a corresponding reduction in damage, pilferage, and handling costs. The iterative process of assessment/adjustment described in the foregoing results in an initial design configuration baseline which is essentially a preliminary statement of the ship system requirements. Such ship system requirements include the followingfor a merchant vessel : a

a a

a

a

Payload (cargo/passenger capacity and description) Sustained sea speed and endurance Number of containers, holds, refrigerated spaces, etc., for balanced service Limits to overall diiensions such as length, draft, Beam, etc., for operability on required service Loading-discharging methods and capacities Hotel requirements such as heating, ventilation, air conditioning, galley, public spaces, power, and lighting Crew or manning requirements Automation and mechanization objectives Reliability and logistic support objectives Special requirements for navigation and communications Maneuverability (steering, handling, stopping, and backing) Anchoring and mooring

8

MARINE ENGINEERING POWER P

R

W

I

M I S S I O N PROFILE L ~

Main Propulsion System Shaft horsepower Propeller rpm Specific fuel consumption and bunker capacity Space and weight objectives Adaptability to ship configuration

REQUIREMENT A c n v SELECTION OST / EFFECTIVENESS

@QUIP AVAIL & CHUACTERlSllCS

I F € CYCLE COSTS

CONFIGURATION OI SYSTEMS

Auxiliary Ship Systems Power and lighting Steam-galley, deck, and heating systems Heating, ventilation, afid air conditioning Firefighting, bilge, and ballasting Fresh water

CONSUMAILES & tNDURANCE

Fig. 3

3.4

M & R AND LOGlSllCS SUPPORT

Hull Engineering Systems Anchor handling Steering engine and bridge telemetering control Cargo handling gear, such as winching systems, burtoning, and swinging boom Crane systems Bulk cargo systems, self-unloaders, etc. Container systems Palletized systems Tankering systems, such as cargo piping and pumps

Propulsion machinery preliminary design spiral

Speciflc

Marine

Engineering Requirements.

The broad requirements of the ship system as just established must be translated into specific performance capabilities by the naval architect and the marine engineer. Since the requirements established a t this point are broad (e.g., unmanned engine room), subsequent investigations may show that some of the requirements cannot feasibly be fulfilled; in which case, all considerations must again be re-evaluated. Most of the broad requirements of the ship system cannot be analyzed independently of the others; and further refinement of each, to a degree, involves yet another iterative design process which is analogous to a slowly closing spiral that gradually approaches a point of fixation. Figure 3, which was taken from reference [23], is a diagram of, this sort of iterative spiral. The marine engineer utilizes a procedure similar to that indicated by Fig. 3 when performing the design comparisons and trade-off studies required to establish specific design requirements in the area of his cognizance. Such specific design requirements will generally be of the following classifications:

Electronic and Navigation System Commupication, exterior and interior Radar Loran, Decca, RDF, etc., navigational aids Military electronics, sensors, command and control systems, weapons directors, tactical data systems, and electronic countermeasures The procedures which are used when designing the engineering aspects of a ship may best be illustrated by outlining the process of designing a ship from a marine engineering viewpoint. This is done in the following sections and is initiated by a review of the procedures used in developing the main propulsion system requirements.

Developme~~tof Main Propulsion System Requireme~~ts 4.1 Overall Considerations. The basic operating requirement for the main propulsion system is to propel the vessel a t the required sustained sea speed for the range (or endurance) required of the vessel and to provide stopping, backing, and maneuvering capabilities. I n the case of a military vessel, which rarely operates a t its maximum rating, the speed requirement may be partly stated in terms of a mqimum flank or burst speed, which need be sustained for only a short percentage of the operating life of the vessel, in conjunction with a

-

more efficient lower speed for long-range endurance. A further restriction is that the main propulsion system must fulfill all of the basic operating requirements at a cost within that allocated during the preliminary studies of the ship system; otherwise the preliminary studies must be re-evaluated. Many factors must be considered in selecting the main propulsion system. Reliability is of the utmost importance since the safety and security of the vessel will depend upon it. Specific fuel consumption, bunker

GENERAL

cot

capacity, type of fuel required, fuel availability, space and weight requirements, and the adaptability of the propulsion system to the overall ship configuration are closely related to the type of plant selected and must be evaluated. Comparative costs, that is, first costs and operational costs, are also major considerations in tradeoff studies. Before entering into the process of selecting the main propulsion plant, it is necessary that the power required for sustained operation and endurance be tentatively determined. Since the space and weight requirements for the propulsion plant can have a significant effect on the ship configuration, and since the dimensional and form characteristics of the hull and its approximate displacement are required in order to arrive at an estimate of the propulsive power required, it is apparent that the marine engineer must coordinate his activities with the naval architect from the earliest conceptual design stage in an iterative preliminary design process such as that discussed in the previous section and illustrated in the preliminary design spiral, Fig. 3. 4.2 Determination of Ship Resistance. The general subject of ship resistance falls within the domain of naval architecture as opposed to marine engineering. For this reason, a detailed treatment of the subject is left to reference [21]; but for completeness purposes, some of the considerations involved warrant a brief review. The most reliable means of determining the resistance of a ship is to construct a scaled model of the underwater portions of the ship and conduct model resistance tests at one of the towing tank installations. .However, for several reasons such a procedure is far from feasible during the preliminary design phase: one is that sufficient time is not available; another is that the ship dimensions frequently change during the preliminary design phase; and another is that repeated testing would be prohibitively expensive. When tentative values have been established for the ship payload, sustained sea speed, and principal dimensions, an approximate assessment of the ship's resistance aan feasibly be obtained by utilizing the results obtained from a series of tests with systematically varied hull forms. There are principally two such test series: the Taylor's Standard Series [32, 331 and the Series 60 [34]. The Speed and Power of Ships [32], which was the original presentation of the Taylor's Standard Series data, is in tm exceptionally clear and concise form for preliminary design purposes and is a classic that is extensively used by practically all design activities; if not used directly, it la a t least used as a standard for evaluating the relative merits of any particular ship configuration. Although the use of series test data to estimate the resistance of ships is straightforward, the process nevertheless entails a considerable amount of tedious labor. In the event that the accuracy of an estimate is somewhat I&a important than the rapidity with which it can be made, a statistical method similar to that developed by Johnson and Rumble [28] can appropriately be used. Johnson and Rumble developed a simple approximate

statistical method of estimating the weight, displacement, speed, power, and other principal characteristics of a wide variety of dry cargo ships and tankers by averaging plots of a substantial number of actual designs. A number of marine engineering design activities have reduced their data on existing design series to a similar basis such that it is suitable for programming on an electronic computer; this enables approximate investigations of the parametric type to be made rapidly. As noted previously, reference 1211 contains a detailed discussion of the methods which are employed to obtain resistance estimates for ships. 4.3 Selection of the Propulsor. Once the ship speed, requirements and resistance have been tentatively established, it is necessary to select the type of propulsor. With considerations restricted to the type of propulsor for the moment, as indicated by Fig. 4, which was taken from reference [35], some types are inherently more efficient than others for particular applications. The abscissa on Fig. 4 is in terms of the Taylor power coefficient,B,, which is defined as:

where

N = propeller rpm P = power, hp V4 = speed of advance, knots The efficiency of propulsiop devices, including jet propulsion, is presented in a somewhat similar manner in reference [36]. The selection of the propulsor may not be a simple process, particularly in marginal cases, because in order to establish the type of propulsor it may be necessary to a t least tacitly select the type of main propulsion machinery. For example, the gain in efficiency offered by selecting contrarotating propellers versus a Troost B Series propeller (discussed further in the following), for a cargo ship, must be assessed in light of the impact on the main propulsion machinery and shafting arrangements. Similarly, the selection of the number of propellers may be a multifaceted problem. I n general, vessels may be single, twin, triple, or quadruple screw. That is to say, the total power required to propel a vessel may be distributed (usually equally) between one, two, three, or four shafts and propellers. From the point of view of initial and operating costs, fewer numbers of propellers are preferred, but the magnitude of the ship effective horsepower requirements or restraints on the propeller diameter may force a multiple-screw arrangement because of excessive propeller loading and the attendant danger of cavitation associated with unduly small propeller diameters. I n addition, there may be other factors in a given case, such as less vulnerability, more maneuverability, or take-home capability in the case that propeller damage may be likely in service, which favor an arrangement with a larger number of propellers. )

GENERAL CONSIDERATIONS

MARINE ENGINEERING

-

BP

Fig. 4

Cornparim of opfimum ettlckncy valuer

There are several extensive systematic series of fixedpitch propellers which have been model-tested and are in a form convenient for design selection purposes. Of these, probably the most suitable for design approximation is the Troost B Series of three, four, five, six, and seven-bladed propellers although there are others which may be used [21]. I n the usual case, the maximum propeller diameter that will provide adequate propeller submergence for the operating draft of the vessel and provide ample tip clearances as well as adapt to the stern configuration of the vessel so as to minimize propeller blade frequency excitation forces may be used for propeller selection purposes. The propeller design established during the preliminary design phase is generally very close to that obtained from later, morerefined design studies. A trade-off study must be made between the propeller rpm which is required from a maximum propulsive efficiency viewpoint and propeller rpm constraints imposed by prime mover/transmission size, weight, and cost considerations. The propeller rpm which is necessary to achieve a maximum propulsive efficiency is frequently considerably lower than that which is feasible from the viewpoint of the prime mover/transmission (due to the greater torque and hence machinery size associated with lower propeller speeds). Furthermore, attainment of the maximum propulsive efficiency does not necessarily constitute the most cost-effective system. Propeller characteristics are in general such that the propeller can be designed to operate a t an rpm somewhat greater than that corresponding to the maximum propulsive efficiency without incurring a serious efficiency

fv diiemnf Wpcn of propulm

'

penalty. Whiie no significant penalty in efficiency is incurred with propeller rpm's slightly greater than that for peak efficiency, significant savings in the first costs, size, and weight of the prime mover/transmission can be realized due to the lower torque rating (with the power remaining the same). The most cost-effective propeller rpm is selected by conducting a trade-off study which balances the propulsive efficiency against the size, weight, and cost of the prime mover/transmission. 4.4 Establishment of Propulsion Plant Shaff Horsepower Rating. Good practice dictates that a ship's

propulsion plant be rated such that the desired ship speed can be attained with reserve shaft horsepower capabilities. Factors to be considered in establishing the reserve capability include fouling and roughening of the hull, roughening of the working sections of the propeller due to cavitation or erosion, and erosion and deposits on the internal flow passages and working elements of the prime mover and power plant parts; all of which result in a significant performance degradation (approximately 5 to 15 percent) in time. It is also important that the vessel have a reasonable ability to maintain speed in moderately rough seas and adverse weather conditions. The usual practice for providing such a margin is to utiliie the parameter sustained sea speed, which is defined as that speed which is obtained a t some percentage of the installed maximum shaft horsepower, during trials, a t design load draft, under favorable weather conditions, when the vessel and engines are new, and the hull is clean. The percentage (or the so-called service factor) of the maximum shaft horsepower used to establish the sustained sea speed is ordinarily taken to

be 0.80 for cargo ships, which may be continuously loaded during the various legs of a voyage, and 0.90 for tankers, which in general are loaded on the outgoing leg of a voyage and in b a a s t during the return leg. However, depending upon the itinerary, the type of maintenance that is predicated, and mean time between dry docking and overhauls contemplated, the service factor used in a particular case may be somewhat Werent. 4.5 Selection of Main Propulsion Plant. Considerations concerning the selection of tbe main prop h i o n plant cannot be deferred until the propulsor, propulsion plant rating, etc., have been established, which may be suggested by the order of this discussion. Instead, the type of main propulsion plant is generally assumed a t the time the type of propulsor is established. Nevertheless, a final review of the main propuleion plant selected is one of the last tasks accomplished. Selection of a main propulsion plant entails the marrying of a power geeerator/prime mover, a transmission system, a propulsor, other shipboard systems, and the ship's hull. A myriad of possible propulsion plant arrangements may be considered by the marine engineer in making the selection. As indicated in Pig. 5, even when the range of considerations ia confined to the mo8t popular drives for fixed-pitch and controllable-pitch propellers, tbe number of permutations open to the marine engineer is sizable. It may be noted from Fig. 5 (which neglects infrequently used arrangements such as, for instance, directdrive steam turbines or the out-of-date reciprocating steam engine) that in modem ships only large-bore, slow-speed diesel engines are directly connected to the propeller shaft. Transmission devicea such as mechanical speed-reducing gears or electrical generator/motor transmissions are otherwise required to make compatible the relatively high rpm necessary for an economical and small prime mover and the relatively low propeller rpm nece8sary for a high propulsive efficiency. In the case of steam turbines, medium and high-speed diesel engines, and gas turbines, the high rpm inherent in a compact prime mover design and the low speed suited to the marine propeller is reconciled with speed reduction geah. Gear ratios vary from relatively low values for medium-speed diesels up to approximately 50 to 1 for a compact turbine design. An electricd transmission has attractive features, dthough its first cost tepds to be somewhat high; in this owe, the prime mover drives a generator or alkrnator

STEAM TURBINE(S1 n ~ v ~ n LLEICUTS w ~ u ~

WITH

DIESEL ENGINES IEDIUI

8,,ED0n

IRCVCRSI*.l

MECHANICAL

c,,BINATloN

REDUCTION GEAR

A:."S::.lnIvEnwI*.l 4.N.D,%JY,"."ty& -

y;r;~~~;;~~~;;;;~~;~

DIRECT SHAFTCOUPLED

cnA'TrvCE~WtOWpEnpOn*AuC~~

INOM nEvCReIN0)

-

COMBINATION DIESEL ENGINE AND QAS TURBINE DIESEL, UEDIUI S ~ ~ E D O R U I ~ U S ~ E D ,

MECHANICAL REDUCT I O N GEAR1nEvEn.Iu.J

IncvCnsINeI

-

FIXED PROPELLCR

eA,~Un,lNE,Hc,vvOU

""'ICVEm8~*~)

-1

ELECTRIC DRIVE lnEvnn#lu.)

DIESEL ENGINC~S) 'OW ""O

InIVEn#luo)

'

C

GAS TURBINE unrw ourv on UIOU P C R ~ D R ~ A M C E - luo*nEVEnsI~@~

ELECTRIC DRIVE IneVcn8Iue) MECHANICAL REDUCTION GEAR

MECHANICAL REDUCTION GEAR

d--

CONTROLLABLE AND REVERSIBLE PITCH PROPELLER

DIESEL CNGINES .,,,,, 8,EEoon , ,,,,, tao* ~LVIIOIUSI

Fig. 5

Alternatives in the wlection of a main propulsion plant

which in turn drives a propulsion motor having a large number of poles which is either coupled directly to the propeller or drives the propeller through a low-ratio reduction gear. Electrical drives may be either a-c or d-c; an a-c transmission is somewbt favored since it is lighter and cheaper, but it involves special design considerations in order to provide satisfactory maneuvering torque characteristics and becomes more comple~than a d-c transmission especially when the 'prime movers are diesel engines which may be stalled if J o e too abruptly. Reveming may be accomplished by stopping and reverb ing a reversible engine, rts in the case of many reciprocating engines, or by adding reversing elements in the prime mover in the case of steam turbines. It is geperdljl impracticable to provide reversing elements in gas turbines, in which case a reversing capability must be either provided in the transmiwion system or in the propulsor itself. Reversing reduction gears for mch tralismissions are available up to quite subs$antial powers, and controllable and reversible-pitch propellers also have been used with dim1 or gas turbine drives. Electrical drives provide reversing by dynamic braking and ener@zing (plugging) the electric motor in the reverse direction.

Section 5 Main Propulsion Plant Trade-Off Studies 8.1 Fundamental Concepts. The design of the maahinery plant, like many other general design projects, I@y consists of a correlation of a number of units end

elements into a functioning system which gives a desired performance. This entails selecting components, adjusting each to the constraints imposed by all others, and

GENERAL CONSIDERATIONS

MARINE ENGINEERING

12

arranging them so as to achieve the required system performance, a satisfactory configuration, and an equitable life cycle cost. There are a number of design decisions which must be made in formulating a main propulsion plant design. For example, the prime mover must be selected with the major alternatives being a diesel engine, oil-fired steam turbine, nuclear-fueled steam turbine, gas turbine, a combined design, or a special design such as that required for surface-effect vehicles. And once the generic type t been established then the major characterof ~ i a nhas istics of the plant must be selected. Questions which must typically be answered in selecting the major design characteristics of the propulsion plant are: Should a &&el plant be high speed, medium speed, low speed, two cycle, four cycle, and the like? Or, in connection with a fossil-fueled steam turbine plant, should the boiler have natural circulation, forced circulation, or no recirculation at all (once-through type)? With gas turbine installations, there is the choice of simple or regenerative plants. Innovations in nuclear technology continue to provide new alternatives in the design of nuclear plants. The most controversial subject in marine engineering is the relative merits of the various types of main propulsion plants and each type of plant has its own advocates, who often exhibit excessive enthusiasm for their particular type. Since a variety of types is used more or less extensively in a number of ships, it can be c~ncludedfrom this fact alone that all types bave their ~ l a c pand that the only way to determine the most suitable choice of main machinery plant is to consider all of the factors involved in each particular application. The selection of a ship's main propulsion plant may be influenced by previous practice, as is the case with most complicated engineering systems. Ordinarily, pertinent plans and essential data relating to the machmery of other ships, some perhaps rather similar to the one in question, will be available. If this information is aivailable and in a proper form, first approximations can often be made without detailed study, thereby reducing the range and number of variables that must be given detailed consideration in the preliminary stage. There are many factons which should be considered in conducting trade-off studies involving the various types of main propulsion plants; the more important factors are : 1 Reliability 2 Maintainability 3 Space and arrangement requirements 4 Weight requirements 5 Type of fuel required (including fuel treatment) 6 Fuel consumption 7 Fractional power and transient performance 8 Interrelations with auxiljaries 9 Reversing capability 10 Operating personnel 11 Rating limitations 12 Costs

4

I n addition, however, the selection of the type of main propulsion plant can be influenced by intangible personal factors reflecting the backgrofind or personal preferences of those interested in the construction or operation of the vessel, and greatly influenced by the experience of the operating personnel available to them. The aforementioned factor8 will be given a more detailed discussion in turn. 5.2 Reliability. Of all the factors which must be considered in selecting the most suitable type of machinery, reliability in service is one of the most important and should be given proper emphasis. The design effort devoted to this consideration has been receiving increasing emphasis during recent years [37-44]. This has been attributed to the increasing'complexity of the more modern equipment and the increased reliability requirements which are associated with the trend toward reduced manning. Breakdown in the propelling machinery may mean the loss of ship availability (or even the loss of the vessel), which is a very serious matter for the owners and operators. Considerations other than reliability, such as fuel economy, weight, space, and first cost, which may seem to be important in the early stages of the design, later become surprisingly insignificant when compared with irritating and costly service interruptions which can result from inadequate reliability. Accordingly, developmental features should be proven ashore where failures are of little consequence as compared with failures at sea. The method of establishing ratings of the various power plant components should be analyzed for service and design margins so as to ensure the high degree of reliability required for the safety of the vessel. Assurance should be provided that reasonably conservative horsepower ratings are used for design purposes since in some cases there is a tendency for ratings to be stated as that obtained on block tests under ideal laboratory conditions as opposed to the lets-perfect environmental conditions that are encountered in marine service. Evaluating the service and design margins is d i c u l t ; the type of fuels and the pressures, temperatures, and pressure ratios used in the design have a significant effect on the plant reliability. However, realistic trade-off studies require that either the degree of conservatism be consistent between various candidate power plants or an allowance be made for the differences. 5.3 Mai~tainability. Both preventive maintenance and correctiye'maintenance requirements must be considered in selecting the type of machinery to be used in a propulsion plant [&50]. Preventive maintenance has a direct impact on manning levels and operating costs. If the equipment installed requires frequent preventative maintenance, such as greasing, packing, cleaning, and parts replacement, crew personnel must be provided to carry out these duties. This is an important consideration as the cost associated with one crew member over the l i e of a ship is a startling sum, particularly if he must be highly skilled; additionally, the cost of the materials required for preventive maintenance adds to

I I

operating costa and can become significant especially when special tools and equipment are required. Corrective maintenance must also be considered in light of the manning requirements (as regards both manpower and skill level), materials, and tools required. Furthermore, the various modes of equipment failure should be studied in order to identify the failure modes which would adversely affect the propulsion plant operation (the effect could be either in terms of performance degradation, corrective maintenance requirements, downtime, or a combination of these considerations). Failure modes which have unacceptably adverse effects should be further analyzed to identify methods of reducing the likelihood or consequences of their occurrence (e.g., by means of redundancy or selecting other design alternatives). 5.4 Spare and Arrangement Requirements. Some years ago the minimum space required for the machinery plant of a merchant ship was a relatively unimportant consideration due to the tonnage laws in effect a t that time. Formerly, if the actual propelling machinery space exceeded 13,percent of the groas tonnage of the ship, then 32 percent of the gross tonnage of the ship could be deducted in computing the net tonnage, which is the basis for tax assessments, harbor and canal dues, etc. As a result, a special effort was then made to ensure that the space required for the propelling machinery was a t least 13 percent of the gross tonnage of the ship. The tonnage laws have subsequently been modified, however, and such an artificial condition no longer exists. I n most ship desigd configurations, an intensive effort is made to minimize the space required for the propulsion plant. In general, the space required for the machinery space is considered to be deducted from that which can be used for other purposes (e.g., carrying cargo); and a maximum effort is accordingly made to restrain the dimensions of the machinery space. In some ships, such as tankers, this is not as critical a factor. Minimum space requirements are almost impossible to generalize satisfactorily for different types of power plants. There is no substitute for making at least a preliminary ship arrangement layout to determine the effect of the power plant on the overall machinerv mace oonfiguration. In order to illustrate general dff%nces in this respect between principal propulsion plant types, representative machinery arrangements in typical merohant vessels are shown in Figs. 6,. 7,. 8, and 9 for a slteam turbine, diesel, nuclear, and a gas turbine plant respectively. There is a wide range of flexibility in the design of the propulsion plants illustrated; therefok, the oonfigurations shown should only be considered representative. 5.5 Weight Requirements; The importance of the weight of a main propulsion plant varies depending upon the particular application. I n the case of tankers, whose cargo capacity is limited by draft restrictions, the weight of the main propulsion machinery represents oargo foregone. Cargo vessels, on the other hand,

13

seldom operate at their full load draft; furthermore, they have chronic stability problems due in part to the extensive amount of cargo handling gear located high on the ship. As a result, the weight associated with the main propulsion machinery, as such, is mildly advantageous in that it improves the stability of the ship. I n general, naval vessels have chronic weight problems, particularly since the advent of the major emphasis on shock resistance; and shipboard equipment is carefully analyzed from the viewpoint of weight reduction. Representative pmpulsion plant weights (without fuel) are shown in Fig. 10, where the specific weight (the weight of the complete propulsion plant per unit of rated shaft horsepower) is plotted versus shaft horsepower rating. Representative propulsion plant weights, including fuel, versus the plant shaft horsepower rating are shown in Fig. 11. This plot permits a proper comparison to be made between petroleum-fueled plants and nuclear plants; for the latter the weight of fuel is not significant. Propulsion plant weights have been greatly reduced over the years. This trend is expected to continue, particularly as regards nuclear plants, due to the relatively large amount i5f research and development expended on this type of plant. 5.6 Type of Fuel Required. Although solid and gaseous fuels (coal, uranium, and natural gas) play important roles in worldwide energy production, by far the greatest proportion of the fuel buined aboard ships is petroleum fuels. Virtually all petroleum fuels are obtained by fractionating or cracking crude oils obtained from the world's various oil wells. There is a wide spectrum of petroleum fuels from which a choice may be made; some of the more important alternatives are given in Table 2. Table 2 TYPEOF DISTILLATE Light

Petroleum Distillates and Their Uses CLA~SIFICATION Intermediate naphthas Kerosene

Medium Heav

&~du&

Gas oil ~ubricatin~ oils Residual fuel oils Refinery sludges

COMMON UNRESTRICTED USES Aviation gasoline Motor gasoline Tractor fuel Gas turbine fuel Heating fuel Diesel fuel Not used as fuel Boiler fuel Refinery fuel

I n general, oils with higher viscosity are less expensive; however, an additional major consideration js that higher viscosity fuels have greater concentrations of impurities and harmful constituents. The fuel oil selected should be determined on the basis of the lowest overall cost, taking into consideration factors such as initial costs, handling costs, and equipment maintenance costs which can be attributed to the fuel. Factors which must be borne in mind, relative to handling and equipment costs, when selecting a petroleum fuel are fuel constituents, type of metals which will be

MARINE ENGINEERING

GENERAL CONSIDERATIONS

15

I

PLAN VlEW OF MACHINERY SPACE

PLAN VlEW OF MACHINERY SPACE

P L A N VlEW OF MACHINERY SPACE P L A N VlEW OF MACHINERY SPACE

ELEVATION

ELEVATION I BOILER 2. H.P. TURBINE 3. L.P TURBINE 4. REDUCTION GEAR 5. CENTRALCD~~TROLROOM S. STEAM TURBO-GENERATOR 7 WORKSHOP B. CONTAMINATED SThAM GENERATOR 9 DISTILLING PLANT

10. MAIN CONDENSER I I. MAlN CIRCULITING PUMP 12. 13. 14 15. 16. 17.

THRUST BEARIUG FORCED DRAFT FAN STEAM AIR HEATER UPTAKE DEAERATING FEED HEATER LUBE OIL SUMP TANK

Fig. 6 Steam turbine powor pknt

*

I. MAIN ENGINE

2. 3. 4 5. 6. 7. B.

THRUST BEARING MAIN ENGINE CONTROL CONSOLE MAIN SWITCHBOARD TuRBD-GENERATOR DIESEL GENERATOR DONKEY BOILER FUEL OIL PUMP

9 MAIN AIR COMPRESSOR 10 ENGINE I I. MAIN STARTING AIRFUEL TANUHEATER 12. DISTILLING PLANT I 3 LUBE OIL COOLER 14 AIR COMPRESSOR IS. WASTE HEAT BOILER I S EXHAUST SILENCER

Fig. 7 Low-speed diesel power plant

I. NUCLEAR STEAM GENERATOR 2. MAIN BLOWER TURBINE 3 AUX BLOWER 4. H P TURBINE 5. L.P TURBINE

6. 7. 8. S. 10.

REDUCTION GEAR TURBO-GENERATOR MAIN SWITCHBOARD WORKSHOP AIR COMPRESSOR

I I.

AIR TANK

12 13 14. 15. 1s 17. 18 19.

CONTROL CONSOLE DISTILLING PLANT CONTAMINATED STEAM GENERATOR DEMINERALIZER CAUSTIC B ACID STORAGE COMPONENT COOLIN0 SYSTEM LUBE CONDENSER OIL GRAVITY TANK MAIN

22 01

MAIN CIRCULATING THRUST BEARING PUMP

22. 23.

DEAERATING FEED TANK LUBE OIL SUMP TANK

I. MAIN ENGiNE

a. a. 4.

5. 6. 7.

a.

AIR INTAKE PLENUM EXHAUST DUCT COMPRESSOR GAS TURBINE REDUCTION GEAR LUBE OIL SUMP TANK STEAM TURBO-GENERAT~R

h.9

9. M I I N SWITCHBOARD 10. MAlN CONTROL CONSDLE I I. THRUST BEARING 12. PORT USE BOILER 13. DISTILLING PLANTS 14. DIESEL GENERATOR 15. WISTE HEAT BOILER 18. STEAM DRUM

Ggs turbine power plant

Fig. 8 Nuclear power plant

degradation being dependent upon the type of prime mover and its design parameters. It ia extremely important that fuel combustion technology be properly taken into account in any realistic appraisal of propulsion machinery life cycle costs and in the selection of an optimum fuel for a given set of circumstances. Much material has been published on economic oom~arisonsof ~uclearversus fossil fuels for shipboard we. These studies are clouded by the fact that the nuclear technology is subject to strong governmental Influence. The Atomic Energy Commission closely controls the manufacturing of nuclear fuels in the United Btates rigid licensing procedures; however, there several private firms which are engaged in the production of nuclear fuels.

I n the case of very large-capacity central station plants, where the cost of transporting coal is quite important, nuclear fuel has appromhed economic parity with f w i l fuels. However, ship power plants generally fall into a small-capacity category as compared to central station plants; consequently, widespread application of nuclear power in merchant ships will probably await further 'advancements in nuclear reactor practice and technology. Nuolear power for large naval ships is advantageous in that it eliminates the requirement of frequent refuelings, thereby aueenting the shipPs military effectivenew Nuclear power b p&iCularly advantageous in the case of submarines and has pmvided them with new dimensions of operability, submerged endurance, and military effectiveness.

16

GENERAL CONSIDERATIONS

MARINE ENGINEERING

The efficiency of gas turbine cycles is highly dependent upon factors such as the turbine inlet temperatures, the amount of regenerative heating, the pressure ratios, and methods of staging and matching the characteristics of the various compressors and turbines used. These are discussed in Chapter 6. 5.8

SHP RATING OF PROPULSION PLANT (THOUSANDSI

Fig. 12

I

I

201 I I I I 14 1 8 ' 2 2 26 50 34 38 42 SHP RATING OF PROPULSION PLANT (THOUSANDS)

Fig. 1 0 Specific weight of propulsion plants

8

SHP RATING OF PROPULSION PLANT (THOUSANDS)

Fig. 11

,

All-purpose fuel consumption

Weight of cargo ship propulsion machinery plus fuel for a 10,000mile voyage

As indicated in the foregoing, the selection of a fuel n e e success of the ship. An analysis of life cycle costs which fails to take the maintenance factors and other various aspects of the fuel selection into proper consideration would not be expected to be meaningful. 5.7 Fuel Consumption. Differing types of propulsion plants have inherently different thermal efficiencies and specific fuel consumption rates. A heat balance is the fundamental tool used i n determining the fuel consumption associated with a power plant, and it is given a detailed treatment in Chapter 2 for a steam turbine propulsion plant. Heat cycles related to other types of prime movers are discussed, to the extent deemed appropriate for a text of this sort, in the chapter applicable to the type of prime mover under consideration. The fuel consumption chmacteristics of various types

is a multifaceted process which may greatly i

of propulsion plants are expressed by Fig. 12, which illustrates the relationship between fuel consumption and size for the more usual propulsion plant alternatives. The fuel consumption indicated in Fig. 12 includes that required for the main propulsion plant, auxiliaries, and normal hotel loads; no allowance has been made for extraordinary service, such as the hotel load on passenger ships, cargo heating and tank cleaning on tankers, and cargo refrigeration. Figure 12 is not intended to be uaed as a substitute for detailed fuel consumption calculations; it is intended to illustrate only the general characteristics of the propulsion plant alternatives. Once the general type of propulsion plant has been tentatively selected, there are several design characteristics which may be selected to enhance the plant fuel consumption characteristics. For example, with regard to a steam turbine propulsion plant, regenerative feedwater heating using extraction steam or reheating of the steam in the boiler after a portion of expansion work has been extracted in the turbines typifies the methods by which the thermal efficiency of a steam cycle can be improved. In general, trade-off studies are required to determine the most appropriate steam cycle. Trade-off studies could consider such parameters as boiler superheater outlet pressure and temperature, condenser vacuum, main turbine efficiency, number of stages of regenerative feed heating, and selection of extraction points. In addition to the presentation made in Chapter 2, several excellent studies have been conducted and published which deal with the effect of cycle variations on machinery plant performance [51-571. These studies are useful in that they provide a sound basis upon which preliminary decisions can be made. Trade-off studies for the purpose of improving fuel economy should similarly be conducted with gas turbine or diesel propulsion plants. Cycles employing diesel engines tend to have higher thermal effioiencies than those employing steam turbines since the cycle works between greater temperature extremes; nevertheless, the overall efficiency of the total power plant can be improved by the use of waste-heat boilers or exhaust-gas turbines.

-

Fractional Power and Transient Performance.

Except for short periods when leaving or coming into port, most merchant vessels operate a t or near full power. Occasionally, the operating schedules include periods a t reduced speed that may be long enough to require special consideration, but such lowering of speed rarely goes below that corresponding to about one-half power. The case of naval vessels is entirely different. They are designed for high speeds for use on those occasions when speed is of great importance. However, most of the operating life of a naval vessel is spent a t moderate speeds, roughly about 60 percent of the maximum speed. Such cruising speeds require only about 20 percent of the normal power for which the machinery is designed. Good economy a t these low speeds is as important as at maximum speed, because it determines the cruising range of the vessel during many operations. I n high-powered naval vessels, therefore, specid provisions are made for economy at low ppwers. These usually include specially designed turbines (with cruising stages or stage arrangements which can be operated in series a t low powers and in parallel a t high powers), and auxiliary arrangements which are especially designed for economical operation at low powers. I n some instances the service requirements of a ship impose severe demands upon the propulsion plant. For oxample, special-purpose vessels may be required to operate for extended periods of time in an economical aruising mode, whereas upon command they may be mquired to reach maximum power in a matter of seconds, A special propulsion plant such as the combined-dieseland-gas-turbine arrangement described in reference [58] may be required to satisfy demands of this severity. 6.9 Interrelations with Auxiliaries. A considerable number of auxiliaries are required to serve the main tngines and for cargo support, cargo handling, ship kbndling, hotel load, and the like. Since in most instances there is a choice in selecting the type of prime mover for the auxiliary equipment, interrelations between the auxiliary equipment and the main propulsion plant must be considered in order to ensure that the dvorall ship is designed in the most effective mqnner. Auxiliaries can in general be driven by either steam or dectric power; when the main engines are driven by rteam, it may be desirable to also drive equipment such Y generators, pumps, and windlasses by steam. In the @$so of diesel and gas turbine drives, where steam is not Os readily available, electrically driven auxiliaries may be more appropriate. A supply of steam for heating purposes is required on moat vessels; the quantity depends on the type of vessel ~ n the d service for which it is intended. If the vessel is

17

steam driven, the supply is easily taken from the main boilers. For diesel or gas turbine driven ships, a boiler or boilers will have to be provided for that purpose. One economical method of doing this is to utilize the hot exhaust gases from the main engines by passing them through a boiler specially designed for this purpose. Such a boiler may also be provided with an oil burner to make up the deficiency, if any, and to operate in port when the main engines are shut down. I n tankers, where a large steam capacity is required for heating the cargo and rather large quantities of hot water are required for cleaning the cargo tanks, the boilers for steam-driven tankers may be significantly increased for this additional load. If the main propulsion plant is driven by a diesel or gas turbine, one or two large boilers may be required especially for this purpose. As may be seen, interrelations between the main machinery plant and the auxiliary equipment can be an essential consideration in the selection of the main propulsion plant. 5.10 Reversing Capability. The provision of means for stopping and reversing a ship is closely lrelated to the type of prime mover selected. Propulsion plants that utilize reciprocating steam engines, diesel engines, or electric motors present no problem in providing reversing capabilities because such components are intrinsically reversible. Steam turbines and gas turbines, on the other hand, cannot be directly reversed and require special provisions. The common solution with steam turbines is to provide special rows of astern blading in the exhaust end of the turbine (in the low-pressure region); in order to reverse, steam is admitted to the astern blading rather than the ahead blading. The solution with gas turbines is not as simple. It is generally not the practice to provide astern blading in gas turbines; therefore special provisions such as electric drives, reversing reduction gears, or reversible-pitch propellers must be provided. In cases where maneuverability requirements are severe (e.g., dredging vessels, tugboats, vessels which frequently pass through locks), controllable and reversible-pitch propellers may be used in conjunction with other types of prime movers [591. 5.1 1 Operating Personnel. The number and caliber of the personnel required to operate a main propulsion plant may be of major importance. Even though other considerf~tionsof a particular propulsion plant may be attractive, if difficulty is anticipated in obtaining suitable operating ~ersonnel, prudence may dictate that the plant be abandoned in deference to others. In the past, the general adoption of new types of machinery has been retarded as a consequence of this practical cogsideration. Over the years, fewer men have tended toward a seafaring life and as a result the total cost to man ships has risen sharply. An adequate number of highly trained men has not been available for ship manning and, as a result, propulsion plants have become increasingly more automated as a means of reducing the number of operating personnel required (see Chapter 21 for a

1+

MARINE ENGINEERING

years, there yet remains a limit to the size of diesel engine which is considered feasible. On the other hand, the rating of the propulsion plant, as such, does not impose a practical restraint on the size of a steam turbine

fig. 13 Relative imtalled cork of propulsion plank

J j c u ~ i o n of automation and controls). This is an effective means of reducing operating costs and is expe&d to continue. It should however be noted that automated ships will generally require more highly skilled operati~gpersonnel. This, in part, offsets the advantage of fewer personnel. turbine It is often said that the operation of machinery requires less engineeriog or mechanical skill than that required in connection with diesel engin% exceptto the extentthat This ie not entirely board maintenance of the main engine is carried outby to a shipboard personnel on diesel turbine &ips. The shorthigher degree than on treliability of steam turbines is usually considered and the turbine to be slightly better than diesel for short periods, that is, plant can sustain more maintenance of steamturbines can be postponed for short perioda in many instances. Diesel engines cannot be neglected without serious effects,and, flexibility of maintenance policies is not recommended for any typeof power plant, it is possibly less cmcial on the steam plant than the diesel. 5-12 Rating Lim;+dions. There are practical limits the power ranges in which the various which For example, typesof pmpu~sionplants are f-ible. which have been even though the rating of diesel has continued to increase over the installed

The ratings of propulsion machinery tend to be disCrete rather than continuous; consequently an additional rating limitation is imposed. As an example, gas turbine designs have been developed for a limited number of discrete ratings. If a gas turbine were desired with a rating different from those available, the cost associated with the development of such a special design would be pn>hibitive; the same situation exists, although to a ,gomewhat lesser extent due to the larger number of ratings available, with the other types of propulsion plants. 5.13 Costs. The installed cost, which is one of the most important considerations in making trade-off studies, is also the most volatile- Pro~ulsionplant price levels are strongly influenced by factors such as material and labor costs, the similarity of a plant with those previously produced, and ~ a n u f ~ t u r e rexisting 's work backlog, and therefore are subject to fiuctuations which depend on the current status of the industryof Nevertheless, the relative costs of the various plants along with the general relatiomhip of plant size and cost are illustrated in Fig- 13- The data presented in Figs. 10, 11, and 13 were largely taken from references

-

1% 611.

BY reviewing the factors enumerated in the foregoing which should be considered in selecting the type of marine propulsion plant, it may be noted that in every instance the fundamental issue is economics. There are three types of costs to be considered: initial (e.g., installed costs), recurring costs (e.g.9 fuel consumption), and contingency costs (e-g-,most aspects of B~ using a technique such as the present-value concept, the C O S ~ Sto be incurred in the future can be their present value So that all of the costs associated with the various design alternatives can be totaled and compared, in light of their contingencies, in arriving at the most advantageous alternative [62].

Setti011 6 Pnliminary Design Considerations 6.1 Introduction. Before the naval architect can firmly establish the dimensions, form, and charactervalues for the machinery space and istics of a weight, requirements, fuel consumption, and other engineehg quantities must be available to him. However, these quantities are dependent upon the vessel dimensions and form. I n order that the analysis may pmceed, tentative values must be selected initially and subsequently refined as the analysis progresses. Esti-

mates based on sophisticated procedures are warranted during the fomulative ~ h m e of s a design because the rapidly changing characteMcs of the supporting data are not commensurate with the accuracy of the calculation; overall methods of comparison which may involve the use of results from previous parametric studies or systematic ft3milies figuration are adequate and are more Preliminary design procedures differ so

,

GENERAL CONSIDERATIONS

19

one design organization to another that no routine pro- increase in initial pressure to increase the thermal codure can be described for this process. However, cycle efficiency 1 percent; or a 40 deg F increase in Home guides regarding specific methods of establishing temperature will have the same effect. Chapter 2 tho engineering features of a ship can be reviewed. But contains a detailed treatment of thermodynamics and it) order to proceed with s, typical example of further heat engineering considerations. dcsign selection steps, it becomes necessary to make It may be noted that the heat balance calculation is noveral presumptions. First, it is assumed that an well adapted to electronic computer calculation, permitoverall study similar to those described in Sections 2 ting parametric studies to be readily made. However, ~ m d3 has been used to establish the payload and s u 5 in providing component data to the computer, care must tained sea speed required of the vessel or vessels. be taken that it is reliable and accurate as the results will Second, it is assumed that the shaft horsepower required be no better than the data entered. The effect of the of the main propulsion plant has been established as following design variables on the thermal cycle efficiency, outlined in section 4. Lastly, main propulsion plant tempered by practical considerations, would normally trade-off studies, as described in Section 5, are con- be investigated at this point: nidered to have been conducted and, for the purpose of Boiler superheater outlet pressure and temperature this section, that a rather conventional cross-compound Condenser vacuum eared steam turbine propulsion plant has been identified Number of stages of regenerative feed heating and ILN the most advantageous type for the particular vessel best extraction points r~tldservice under consideration. Steam reheating in boiler Like other complicated engineering systems, much of a Main turbine efficiency nhip design is patterned after previous successful Turbogenerator efficiency (condensing versus practice. Ordinarily, pertinent plans of other ships, noncondensing) Nome perhaps rather similar to the one under consideraExhaust heat recovery from boiler stack gases Oio11,would be available. Also, essential data relating to Motor-driven versus steam-driven feed pumps Illlosevessels and important particulars of the machinery and auxiliaries ad auxiliaries, their characteristics, and their ratings Utilization of and balancing out of excess auxiliary would normally be available. If this information is exhaust steam properly compiled, it is often possible to make useful Desuperheated steam service requirements Arnt approximations without detailed study and thus Distillers, steam-air heaters, etc. reduce the range and number of variables that must be &on detailed study-to optimize a ship design. Of the foregoing design variables, the largest direct some of the more salient considerations in establishing gain in efficiency will come from increasing the boiler tho design of an engineering plant for a ship, in addition superheater outlet temperature and the boiler to the main propulsion plant trade-off studies described There are, however, several factors which cannot be 111 Roction 5, are reviewed in the following paragraphs. ignored; boiler design pressure must be increased in 6.2 Propulsion Plant Steam Cycle. The propulsion proper proportion with the temperature in order to plr~uthas been established to be of the steam turbine ensure that the turbine condition line does not lead to typo; however, the precise steam conditions and cycle excessive moisture in the exhaust end of the low~rrbngementwould warrant yet another review. The pressure turbine as an erosion problem could otherwise h e ~ tbalance calculation is the basic analysis tool for result. Furthermore, inerewing the boiler delurmining the effect of various steam cycles on the outlet temperature and the boiler efficiency beyond tharmal efficiency of the plant. Standard practices certain limits both lead to costly increases in either the atrd allowances which are recommended in the prepara- boiler design or its mainhnance, or both, which must be l ~ n l rof heat balances have been promulgated by the taken into account. When burning Bunker C residual ~ l l l p '~~a c h i n e r yCommittee of the Society and are fuel oil, eutectic combinations of oxides of vanadium, available in ~ e c h n i c a l kResearch Publication No. 3-1 1. sodium, and potassium can c a w slaggng and accelerated 111 tho absence of specSc component efficiencies and erosion of tubing at relatively low metal surface temurvioing allowances during preliminary design, the peratures. Thus if low-grade residual fuel is to be used, r@UXIlmendationsof this publication are most helpful. it must either be treated aboard ship, or the boiler must Many excellent parametric studies have been con- be specially designed to limit the metallic wall tempersduotml by various design agencies and several have been tures of the superheater tubes and supports; additionally, ~ublinhed[51, 52, 531 which cover the effect of steam the boiler should be designed for ready acceas into the @~adltions and cycle variations on machinery plant per- superheater for mechanical de-slagging, cleaning, and f@lmalce. These may be used as a guide during initial tube replacement. *l@otion and thus minimize the amount of detailed work Another factor which should give rise a cautious ah& must be carried out later during the more refined approach in moving to higher design initial pressures @@WO the design- Fmm parametric studies of this and temperatures is the increasing cost and difficulty in Rnl'tt ollc can derive some useful yardsticks for design assuring the safety and longevity of steam piping, n@lailiOn, such as, for example, that it takes an 85-psig joints, valves, fittings, manifolds, and pressure bound-

MARINE ENGINEERING

GENERAL CONSIDERATIONS

21

GENERAL CONSIDERATIONS

23

MARINE ENGINEERING

desisns vary widely depending upon the type of cargo handled [6&75]; however, some of the more common types are as follows: Winching system, burtoning or swinging booms for dry cargo, i-e., break-bulk cargo or palletized cargo systems Cargo crane systems, either shipboard or onehore ~ u l cargo k systems, such as self-unloadem utilizing either standardized Container containers which lift On/& or standard truck trailers which roll on/off B~~~~systems, utilizing hrges which either lift on/off or float on/off systems, utilizing cargo ~ i ~ ~ i tankering d piping, pumps, and so forth Barge raftlngsystems, ut&ing pushboats or

certain limik), radar su~eillance and warning collision hazards (also within cedain limits), data monitoring and recording of principal voyage data, weather reporting, sounding, and fire detection. Some of the facets of navigation which do not appear to be readily adaptable to automation are: docking and undocking; piloting in nanow channels) harbors, Or , territorial waters where local knowledge is emntial; planning and laying-out of best course and speed, taking into account all potential factors; decisions on slowing Or proceeding with due caution in poor visibility ;maneuvering to prevent collision, determination of safe sea speed, determination of best fix from position fixes, and judgment as to when to post lookouts in foul weatherEngine room control stations appear perfectly feasible to permit the handling of even complex plants by a single licensed officer. The gas turbine and the diesel

The relative of the in selecting the types of cargo handlhg gear, such as the winch desip aeociated with different rigging schemes, hydrau]$ally operated hatch covers, special types of cranes, elevators, conveyors, and cargo pumping systems, should be given a rigorous analysis during the preliminary desisn shge. Close cooperation between the naval archit& and the marine engineer is essential in such and power requirements analyees. The space, be estimated very early in the design of a ship as they may have an important impact on the deck arrangement, the size of the electrical generating plant, and indeed the configuration of the vessel itself. Ca%o refrigeration, cargo hesting, ballasting requiremenh related to cargo handling, buttemorthing, etc., are all imporbnt servke load factors which may result in peak loads not only on the electrical generating plant, but also on the main machinery plant. They must then be included in the design heat balances and electric load anslyses from the emliest stage of the design selection process. For a detailed discu$sion of dry, bulk, and liquid cargo handling systems, see Chapter 16. 6.17 Autorntion rnnd Mechanization. Automation and rnechaoieation of shipboard processes are important means of improving the efficiency of ship operation. These are subjects that are particularly well suited to system engineering analyses in that the cost of developmerit, manufacture, installation, and maintenance of such mechanized or automated equipment is readily compared to the cost of hand labor. However, close sight must be kept on the degree of reliability of autornation where it involves the safety and security of the vessel; furthemore, the training and adjustment of maritime labor to new conditions of operation must be rnnsidered in addition to simple engineering feasibility. There is potential for reducing the burden of bridge duty and reducing the number of operating personnel required for the saf. navigation of the ship by the intraduction of rnonitonng and control devices in a bridge coneole. Some of the facets of navigation which are adaptable to automation and semi-automation are: course steering, dead reckoning, position-fixing (within

engine appear particulady well adapted to automation becsuse of the Simplicity of their control- However, even the steamship with its more complicated plant has been automated to a surprbing degree and developments (see Chapter 21). in this direction continue to be 6.18 Dynamic Effecfso Dynamic effects, principally mechanical vibration but also noise and shock resistance, must be an integral aspect of the preliminary design process as the dynamic cha~acteristicsof the ship and the dynamic requirements for equipment am largely established during the preliminary design stages* The objective is to develop the design so that the desired dynamic Characteristicscan be achieved in an effective manner. Reafisticall~ conceived requirements with regard to dynamic effects require careful and adequate planning during the preliminary design stages in that they may be met without excessive dimculty or undue expense. especially important insofaras Vibration analyses

..

the design of the pmpulsion shafting system is concerned, and particularly its relationship to the excitation forces resulting from the propeller operating in a nonuniform wake. Propeller exciting forces are diicussed in detail in reference 1761 and main propulsion shafting systems in Chapter 11 of this text. As may be noted in the latter, the main propulsion shafting can vibrate in longitudinal, torsional) and lateral modes. Each mode of vibration must be dealt with during the early stages of design. Modes of vibration of the ship's hull as a whole (i.e., as a free-free beam) are discussed in reference 1771. T h m may be vertical, horizontal, torsional, or longtudinal and may occur separately or, in rare case*, coupled. The calculation and re diction of the hull vibration modes is quite complex since the hull girder is far from a Simple homogeneous beam. Hull vibration of this type, may be excited by s~nchronirationwith periodic harmonics of the ~ r o ~ e l l forces er acting either through the shafting, by the ~mpellerforce field interl acting with the hull afterbody, or both. ~ u lvibration may also be set up by unbalanced harmonic forces from the main machinery, and in some cases by impact excitation from slamming or ~eriodicwave encounter.

GENERAL CONSIDERATIONS

Tab* ITEM

Mg~lltlllmrated power

'r'ho(+(Lsteam condition ( !otldnnmr vacuum

listof Machinery for RATING

One set, cross compound, with astern element located in exh u tcasing end of low-pressure turbine 24,000 at 105 rpm 850 F 28.5 in. H 925 g 8t maximum rated power

a

24,000~~h,, cargo

ITEM Line Shaft Beanng8

gbr Diameter

Length Materid Stem Tube Bean'ng Type Length

RATING 8

Re laceable shell, ring oiled 21& in. In. 32%

Caet atex?] pedestal, cover and Oil lubricated 27 in. forward bearing 54 in. aft bearing Ductile iron and babbitt

Vertical, walk-in, five pas4 convection with automatic superheat control by desuperheater Coil in steam drum

20,000 Ib/hr from 875 p i g , 930 F to 775 psig, 575 F

345,000lb at 24,000 shp and 105 rpm

Aft of low-speed gear cssing

27% in.

2 4 f t 11Xin. Solid forged steel, ABS Gr. 2

Superheat Control De.guper& Descr~pt~on Coil in steam drum, steam Burnera after third superheater pms Number her boiler 3 Type Wide range ateam atomking

Air quantity, cfm Air temp in-out, deg F Stm Pr=.-temp, pslgdeg F Air press, drop, in. H ~ O

Ruting

116%

23,500

10048.3

29,400 100-275

62-453

62-453

0.6

1.0

GENERAL CONSIDERATIONS

MARIYE ENGINEERING Table 3 (continued) RATING

RATING

ITEM

~ l ~ ~ t ~ - m e c h adeck n i dmounted , 50 hp, 650 rpm, 230 volt d-c

20-ton cargo hoist 1 ~ t o cargo n hoist

14,500 lb at 105 fpm 18,000 lb at 85 fpm 8,800 lb at 185 fpm 14,200 lb at 116 fpm 8,800 lti at 85 fpm 1250 ft of 76 in. wire rope 800 f t of M in. wire rope

eretors

Drum storwe

Section 7 Specifications

,

31

MARINE ENGINEERING

GENkkAl CONSIDERA'I'IONS

Section 8 Final Design and Working Plans

16

Lifesaving Equip-

65 . Air Conditioningand

33

MARINE ENGINEERING

GENERAL CONSIDERATIONS

of contracts for ordinary merchant vessels where the plans must be developed in a short time. Where oOmposite Plans are not made, the elimination of

interferences and the treatment of wstems in accordance with their relative importance must be accomplished by the cooperation of the various design groups iivolved.

Section 9 The design and construction of a ship is culminated by Sea trials are conducted as a means of demonstrating broad array of tests which demonstrate that the ship is the adequacy and perfomance of those aspectsof a ship in accordance with contract requirements. At the lower which cannot be realistically tested at dockside. sea the test spectrum are those of a q ~ a l i @ - C ~ n t r ~trials l are bmadly classified into twogroups; namely, "ature which are conducted to ensure conformance of machinery trials and maneuvering trials. ~h~ former lnaterial properties to specified requirements, soundness deals with the mechanical and economical performance of cmtings, dimensional accuracy, and the like. Tests of the boders, the proeelling machinery and their nuch these are not Peculiar to marine equipment and auxiliaries, and tests of evapowtors and distillers, Ihu standard quality-control Practices of the manu- together with the anchor and steeringgear and Iaaturer Or are generally relied upon to other equipment which cannot be tested uader actual nrluure the adequacy of equipment in this regard. conditions at the dock. The latter involves calibration Shop and installation tests include those tests which of navigating equipment, the of the n o m a l l ~conducted in the shop after assembly or in ship, and the speed-power characteristics of the ship. tho ~esselat dockside after the installation of the Tests typically conducted during sea trials are as equipment or system to be tested is substantially follows: aamplete. These tests are conducted to prove correct Calibration of navigating equipment rflnombly and proper installation and to demonstrate Speed-power-rpm standardization tests that control and safety devices are functional and properly adjusted. References [81, 821 contain general Economy power teats Full-power endurance tests guidelines which may be used in connection with shop Ahead steej n g tests ihd installation tests for merchant ships; similar, Quick reversal astern and head reach although generally more exhaustive, test requirements Astern endurance tests I0r naval ships are invoked in the specifications prepared Astern steering tests for eech particular ship. Quick reversal ahead and stern reach Anchor windlass tests which strength is a major concern. 8.8 Electric Plant. The procedure for the final '*lo COnektion Of Detail pian'' The design of the electrical installation roughly parallels that A careful review is made of of a ship are made by a large number of for the other working simultaneously in several drafting departmentsthe probable electrical loads and the selected number The administration and practice of the dr*ting organizaand rating of ship,s service generators and emergency tion must aim at complete elimination of physical generators. Vendors, pmposals are obtained and r e interferences between various parts and at a design in viewed for correlation with the general design. which each element is treated in acc~rdancewith its The airing plans for power, lighting, and interior relativeimportance. W r e x a m ~ l e , a P o o r l e ~ o f v e n t ~ ~ ~ commu~cations mnsist of single line diapams and tion duct 4ould not be accepted merely because a deck arrangement plans. The single line diagrams in elementary form, the electrical interconnection perfect lead for a freshwater line or an electric cable is of the various parts of each system. The diagrams desired' sections of the of the cables and c o n d u c t o ~ It is Customary to make, for show the approximate machinery spaces, composite layouts showing everyalong the ship and through the decks. The deck thing in those Spaces; i.e., structure, machinery, arrangements show the wiring on each deck and the These may be to a correct location of all appliances, fixtures and fittings, Piping, ventilation, and scale larger than the ~ s u a arrangement l plans; and develop including radio and navigation equipment. possible interferences' "lVe to ing these wiring plans, consideration is given to carrying Or other large Occasionally, in the case of capacities and voltage drops, directness and simplicity of leads, protection, support, and accessibility. important vessels, such composite layouts are made cf 8.9 null Machinew. The marine engineer is usually practically all machinery spaces This procedure is, concerned with the deck machinery and other mechanical however, slow and costly and cannot be afforded in the

36

MARINE ENGINEERING

GENERAL CONSIDERATIONS

37

An Anal~si*' Naval Engineers Journal, 64 D. M. Mack-Florist and R. H~~~~~~~ dlAn Economio February 1965. Feasibility Study of U n i w States Bulk Carriers, 49 A' J' Ruffini~ standard Navy Maintenance Marine Technology,vol. 3, no. 2, ~ ~ r1966, i l and Management system (3-M System),JJ 65 W. j . Dormm, 'dcombimtion Bulk ,, Bureau Of ships Association of senior Engineers, March Marine Technology,vo~.3, no. 4, October 1966. 66 A. W. Feck andTankem J. 0. Sommerhalder, 'Cargo,, 50 A. Goldman and T. B. Slattery, Maintainability: pumping in M~~~~ and Bulk Carriers, A Majw of SYskm Efectiveness, John Wiley & ~ ~T r ~ i 4, no. ~ ~ July, 1967. ~ ~ ~ Sons, New York, 1964. 67 Leslie A. Harlander, "Further Developmenh of a 51 W' Giblon and Cheater W' "Effect container System1961. for the West Coast-Hawaiian T ~ ~ Of Conditions and Cycle Arrangement on Marine Trans. Power-P1ant Performance as ~eterminedby the Elec68 James J. Henry and Henry J. Kamch, ,,Container tronic Computer," Trans. SNAME, 1961. 52 H- M. Cheng and C. E. Dart, "Cycle and Ships, " Trans. SNAME, 1966, 69 5'. G- EbelJ "An Analysis of Shipboard cargo Economic Studies for a 25,000-Maximum-S~pSteam Power Plant for Singlescrew Tanker InstsllationJ Cranes, " Trans. SNAME, 1958. Trans. SNAME, 1958. 70 E. Scott Dillon, Francis G. Ebel, and Andrew R. 53 M. L. Ireland, Jr., H. W. &marJ and N. L. Goobeck, "Ship Design for Improved Cargo Handling, Trans. SNAMEJ 1962. Mochel, "Higher Steam Conditions for ShipsJ 0hiner3'JJJ paper presented to the International Con71 John F. Meissner, "World Development and foreace of Naval Architects and Marine Engineers, 1951. Movement of Iron Ore, Trans. SNAME, 1962. 54 W. L. Coventry, "Fundamentah of Steam 72 -Harry Benfod, Kent C. Thorntan, and E. B. Turbine The-odynami~s,' Trans. Institute of Ma* Williams, "Current Trends in the Design of rron-ore Bngineers, 1962. Ships, " Trans. SNAME, 1962. JJ

JJ

JJ

JJ

Trans. SNAME, 1965.

*

THERMODYNAMICS AND HEAT ENGINEERING

C H A P T E R II

-O,,,(HEAT

TRANSFERREDI

1.1 Basic Equations. The applied thermodynamics problems of marine engineering depend on the conserve tion of mass and the conservation of energy. The first of these is conveniently expressed by the sional steady-flow continuity equation

h?

+ 9+

Q1.2

=

4

h t i-iwt1,2

Typical applications of the general energy equation occur where the working floid is being heated without work being done (a heat exchanger), where work is being done under adihbatic conditions (turbi~le wheel), W = AC/v ('1 or +here mechanical energy is being degraded under adiabatic conditions and without work being done (flow against friction). The equations that apply in tbese situations are esaily found by eliminating the inapproA = flow area, sq ft priate terms from equation (4). An application is C = flow velocity, fps illustrated by Fig. 1. v = specific volume of the fluid, cu ft/lb Evaluation of the general energy equation usually W s flow rate, lb/sec requires assistance from other equations. The conThe second is conveniently expressed for the usual shady tinuity equation is one. Equations of state for the one-dimensional situationby the general energy equation fluid involved are also frequently needed. The simplest form is the familiar perfect gas equation

1

+

Enemy equation as applied la a single-stage turbine

Typical values of R are 53.34 for dry air; 53.5 for wet air (40 percent humidity, 100 F); and 50.3 for flue gm (15 percent excess wet air and standard fuel oil)For perfect gases, the following state relations also hold :

-

I Numbem

in brackete designate References at end of ckpter.

(10)

Carbon. . . . . . . . . . . . . . Hydrogen. . . . . . . . . . . sulfllr... . . . . . . . . . . . . Oxygen. . . . . . . . . . . . . Nitrogen ... . . .. . . .. . Free moisture. . . . . . . .

0.8775 0.1050 0.0120 0.0040 0.0015

-

Charts PI, must be used. An alternative, particularly adaptable to turbine design work when calculations are 1.0000 made by Computer, is to use the equhons from which Other properties of flue w, such as its viacasity and these tabulations are made. thermal are also needed, and are given in Special relations for steam that are useful in nozzle Fig. 4. conductivity, values for steamand air can be found in the design are the equation of state Steam Tables [ll and Gas Tables [2], respectively. pv = 1.222 (h - 823)' 1.2 Heat Transfer. An investigation of the & * (') term in equations (2) or (4) entails a consideration of the and the equation for isentropic expansion principles of heat transfer. The transfer takes place by molecular diffusion between bodies in contact, or by pl.s = constant electromagnetic radiation between separated bodies. The following two are the corresponding relations for Diffusion between solids is c d e d menone the wet region or both of the bodies are fluids, conduction is nearly

T = absolute temperature, deg R R = a constant characteristic of a particular gsa P = pressure, psf J = mechanical equivalent of heat = 778 ft-lb/Btu g = gravitational constant = 32.17 ft/sec2 z = height above an arbitrary datum, ft Q = heat transferred, Btu/lb W r = external work done, Btu/lb

p0.07v= 0.467 (h - 366) p ~ . ~= s constant

(11) Units are psi for p, cu ft/lb for v, and Btu/lb for h. These equations are for use only in the vicinity of normal turbine state h e s , and not for use at high superheat with low Pressure, with very wet steam, or in the reheat region. In boiler design work, the sensible heat, and specific heat of the flue gas must be known. These are presented in Figs. 2 and 3 for a standard grade 6 or residual fuel oil of the composition (by weight) tabulated below when burned in air with a 40 percent relative humidity at a temperature of 100 F.

J. B. Woodwad, III

Review of Fundamentals

39

Re. 2 Selulbk heat of gases

MARINE ENGINEERING

THERMODYNAMICS AND HEAT ENGINEERING

where the subscripts o and i designate the outside and inside surfaces of the tube. Heat transfer problems frequently involve conduction through successive layers of distinctly different conductivity. Formulas for this type of problem are readily derived, as are formulas for the transfer of heat through cylindric composite walls. b. Convection. The convective heat transfer between a fluid at a largely constant bulk tempe~atu~e TB and a surface at temperature T is expressed by

Q = h j # ( T ~- Ts)

property of the material conducting the heat. It is generally a function of temperature, particularly for liquids snd gases, but the effect of temperature is sufficiently weak that conductivity can be treated as a constant in most problems. Fourier's Law can be expressed for one-dimensional problem as

(16)

where hj is the film coeficient of convective heat transfer. major practical problem in applying equation (16) in the evaluation of the fdm coefficient for the several distinct mechanisms of flow and thermal behavior possible in the fluid. Single-phase convection occurs when the fluid involved uoither boils nor condenses at the solid surface. Familiar axamples abound aboard ship; for instance the water side of condenser tubes, both sides of the tubes in liquid-toliquid heat exchangers such as lube oil coolers, and the gtM side of convective heating surface in boilers are typical locations where this mechanism is prominent. Tho value of h, is generally a function of fluid properties, of the fluid' velocity, and of its degree of turbulence. Under conditions existing in a typical condenser tube, far example, the value of h is likely to be in the neighborhood of 1000 Btu/hr-sq ft-deg F, while on the gas side of r boiler tube, the value of h j can be 10 Btu/hrmsq ft-

tho situation is described aa forced convection. When

k

= conductivity, Btu-ft/hr-sq ft-deg F' = conducting area, sq f t

either dropwise or film condensation. The names are quite descriptive of the processes. The rate of heat transfer is much higher for dropwise condensation, and is comparable to that for nucleate boiling, since the drops quickly fall off as they form and thereby expose the surface to more vapor. In film condensation, the condensed film tends to cling evenly to the surface, and so forms a barrier between the surface and the vapor. C. Radiation. All matter emits radiation of one or more kinds. The thermal radiation of practical concern requires only that the matter be at a temperature above absolute zero, and so is characteristic of all bodies. Thp radiation is electromagnetic, and at industrial temperatures lies within the infrared part of the electromagnetic spectrum; but the wavelength is a function of temperature, and at higher temperatures it falls within the range of visible light. The radiation is not, however, monochromatic. A curve of its intensity, IA, against wavelength, A, shows a considerable spread with a peak intensity at a wavelength that is a function of temperature. The total energy emitted is thus the integral of IAover all wavelengths. For a black-body radiator, i.e., one that emits at the maximum intensity at all wavelengths, the integration produces the Stefan-Boltzmann relation (17) for T in degrees Rankine and Eb in Btu/sq ft-hr. But actual bodies are not black-body radiators, and their degree of imperfection must be accounted for by equation (I7) becomes

[A] 4

E

= 1730.

Bodies for which this equation holds are said to be g ~ e g

k is constant, this equation can be integrated for a slab of thickness x, having a temperature difference between faces of TI - Tal to obtain

~f

If the conducting body is circular, as when heat is transferred through tube walls, equation (12)is modified

always -ly modified by the transport of heat by where r is the radius dimension, and fluid in motion; this phenomenon called convection. tube. Integration of equation a. Conduction. Conduction follows Fourier's Law1 which states that heat is diffused at a rate proportional to the temperature gradient; the factor of proportiond t y is known the the~malconductwity, and is a

density differencescaused by expansion or contraction of the fluid near the surface are the principal source of the driving force, the situation is described as natu~alcon-

41

is the length of the

ges

MARINE ENGINEERING

-TI

I n pract.ice, heat exchange by radiation occurs between bodies of different temperatures and different emissivities. The situation is complex because the geometrical arrangements and sizes of the bodies are significant. For an elementary case of two parallel infinite planes, and of respective temperatures and emissivities Ti, €1, T2, e2, the net energy exchange rate is

I

THERMODYNAMICS AND HEAT ENGINEERING

--

-11-

C

COUNTERFLOW

-1

1

-

l"b

PARALLEL FLOW

For a sphere or cylinder, enclosing a smaller sphere or cylinder, the equation is Fig. 5

Simple counterflow and parallel-flow heat exchangers

to the heat that it receives or rejects. If the fluids on both the hot and cold sides of the heat exchanger undergo a change of state, their respective temperatures are constant, and equation (21) applies without change if S is understood to mean the 'total heat transfer area. On the other hand, if there are temperature changes, the temperature difference in equation (21) is not constant throughout the heat exchanger, and in consequence this equation must be integrated for application to the entire apparatus. The case where there is no change of state is illustrated by a simple concentric-pipe heat exchanger, Fig. 5, in which the two fluids flow either in the same direction (parallel flow) or in opposite directions (counterflow). The temperatures of the two fluids are plotted as a function of position for both exchangers. Such a temperature differenceintegrated over the length of the heat exchanger produces a mean temperature difference;because of its logarithmic term it is familiarly known as the log mean temperature difference. I n the

43

general case, the log mean temperature difference can be written as AT,

=

- ATmin AT, log. ATmin

AT,,,,

(23)

Equation (23) is the general expression for AT, for both simple counterflow and parallel-flow exchangers. I n condensers, boilers, and feed heaters, to list several prominent examples, where a change of 'state rather than a temperature change occurs on one side of the tube wall, a derivation of the log mean temperature difference again produces equation (23). If the heat exchanger is multipass, equation (23) must be modified (see Section 2.1 of Chapter 14). I n any case, equation (21)) when applied to the heat exchanger as a whole, is written as

9 = USAT,

(24)

contributions will be additive. Thus, starting at the I n boiler tube banks where the heat transfer fluid is a radiating gas, heat transfer simultaneously Occurs by both radiation and convection. Under these conditions (in order for the two heat transfer coefficients to be directly additive) it is often convenient to express the radiation heat transfer in the form of the artificial heat transfer coefficient

Section 2 Heat Transfer in Boilers where hrl, hj2 = convective surface coefficients at tube out-

T,, ti = metal temperatures at tube outside and inside surfaces respectively k = conductivity of the tube wall X, = equivalent thickness for the circular tube

where the subscript G refers to the radiating gas and 8 refers to the tube surface (see Subsection 2.3 for further discussion on this subject). d. Overall Heat Transfer Coefficient. The typical occurrence of heat transfer in power plant apparatus is cold fluid through an inter- Addition eliminates intermediate temperatures, do-g between a hot fluid and a vening tube wall. Convection and radiation are in- assessment of U as volved at the inner and outer surfaces, and conduction is involved within the tube metal. The rate of heat flow is summarized succinctly by $ = US(T - t) (21) BYa similar pmoess, U can be written for any number of layers. The practical pmcess of heat exchanger design is where T and t are the bulk temperatures of the two fluids, S is the surface area, and U is the ooerd heat transfer often aided by ern~irioalformulas for U which $ve coefident. U is the net effect of the conduction, con- results of suffcient accuracy for industrial purposes. vection, and radiation contributions. To illustrate the They are usually ~ r o m u l ~ a t e dby man~fa~turer's ~1 makeup of U, consider the transfer of heat from a hot associations to standardize methods of ~ a l ~ u l a & and are found in publications such as references [Bland [71. clean tube to a second fluid inside the tube. gss outside As a preliminary, note that the artificial radiation e. Log Mean Temperame Difference. The fluid coefficient ic, of the same dimensions as the convection flowing through a heat. exchanger undergoes either a coefficient hj, is used so that the radiation and convection change in temperature or a change in state in response

[(&)( (&)I + + [ + ++

2.1 Types of Heat Transfer in Boilers. A boiler may be divided functionally into four parts: first, a ~~~OSCFEFA USw(T8 - Tc) chemical reaction chamber where the chemical heat of LHV q~ (to - to)CpR fuel combustion k released and the reaction controlled; = WF(R 1) 1 R second, a steam generating section where heat is transferred to the tubes by radiation, convection, and con(25) duction; th,ird, a superheater, where the steam is super- where heated to the desired degree; and fourth, a heat recovery U = convection heat transfer coefficient section, employing air heaters and/or econombers T c = furnice surface temperature where some of the remaining heat in the flue gas is exTB = furnace exit temperature T F = effective flame radiating temperature S w = convection surface area

where an overall U is estimated and an exit temperature from each bank of tubes is calcdated. The designer must h t estimate the performance of the furnace and 2-2 Heat Tmnsfqr in Boiler Furnaces. Furnace heat transfer is principally radiation, and it is possible to b p t the basic methods of Hottel in reference [5] to evaluate a tohl emissivity in terms of furnace conditions. The problem consists of equating the heat given up by the omb bust ion gases to the heat transferred by radiaflon and convection to the f b a c e surfaces. The

q F = sensible heat of fuel above to sensible heat of gas above to C, = average specific heat of combustion air R = air-fuel ratio FA= arrangement factor FE= emissivity factor

Q T = ~

The heat given up by the gas is evduated by ordinary s t o i c ~ i o m e t ~ means c and the use of a set of sensible heat cumes (fig. 2). The shape emissivity factor, FEFA,has been treated by Hottel, and if the flame fills the furnace, it has been demonstrated that

,

MARINE ENGINEERING

44

0

FLAME EMISSIVITY, EF Fig. 6 Shape emiuivity factor versus Aame emiuivify fw various valuer of cooled surface to cooled surface plus refractory surface ratio (Sc/Srl

THERMODYNAMICS

by the chemical breakdown of the fuel to basic constituents. The flame mass then consists of a cloud of flaming fuel, carbon, some ash particles, and molecules of carbon dioxide, water vapor, sulfur dioxide, oxygen, and nitrogen. Of these constituents, the fuel, carbon, and ash particles and the carbon dioxide, water vapor, and sulfur dioxide molecules radiate. The gas molecules radiate only in certain wavelengths, that &, they are not grey. The solid particles radiate in all wavelengths. These radiations are superimposed upon each other, resulting in an overall radiation which is essentiay grey in character, and the resulting emissivity is independent of temperatufe. Combustion of oil is not instanta~eous,especially when residual oils are fired. The oil droplet first ignites, then burns and breaks down into carbon and hydrogen. The carbon appears aa minute flecks. These small particles make up most of the radiation. Their concentration is a function of burning time, and of the rate of flow of the gases through'the furnace. An expression derived for cp by applying probability theory is

"

TRANSVERSE TUBE DIP,PITCH 0

2

4

6

8

10

12

14

18

18

20

22

24

FIRING DENSITY, WFIPF VF. LB/FT~-HR-ATM

Fig. 8 Wectiveneo factors f a water walls bared on ma1projected arm

Fig. 7 Furnace concentration factor

(26) where fa

= emissivity of a cloud of i n f i ~ t ethickness,

assumed to be 0.95 (27)

where ec = emissivity of the heat absorbing surface e p = flame emissivity

S B = refractory surface area S c = cooled surface area F R C= a geometric factor, dependent on the extent of cooled surface

An ?ppmximation of FRCto a reasonable degree of accuracy is SR when 0 < -- < 0.5 (28) FRC = g Sc

'

+

where ST = S R SC Faired intermediate values may be taken between the two sets of limits quoted, as illustrated by Fig. 6. The radiating temperature T p may be approximated by T p= ( T A ' T E ) ~ ~ ~ (30) where TA' = adiabatic flame temperature with 100 percent theoretical air. Evaluation of the flame total emissivity presents a complex problem. The flame cloud consists of droplets of fuel from the burner nosde which in turn are reduced to smaller fragments by various air and gas currents and

P = furnam pressure, atmospheres L = mean radiating path length; for ordinary marine furnaces, L = 0 . 6 m furnace volume, cu f t an empirical concentration factor, a function of a time parameter W p/PpVp qith WP representing the pounds of fuel burned per hour. Wp/P;Vp is a crude measure of article life but better data on the flame path is lacking. The concentration factor, K, is evaluated from test results on various boilers and plotted against the firing density WP/PPVF, as on Fig. 7. This plot was calculated from the test results on five different boilers, all burning residual fuels. The curve shown represents an average of the test results with 10 to 20 percent excess air. It is necessary dso to consider the question of effective cooled surface. A water wall consisting of tangent tubes may be treated as a surface having an area equal to the projected area of the surface. If the tubes are widely spaced, exposing the refractory surface behind the tubes, the simple projected area of the tubes is not sufficient since the refractory receives some of the direct radiation from the surface and returns only a portion of this heat to the furnace; the remainder goes to the tubes. The effective radiant heat absorbing surface (RHAS) may be calculated by multiplying the projected area of the walls, including backing refractory, by an arrangement factor from Fig. 8, for each area making up the furnace envelope.

Solution of equation (25) is best accomplished by trial-and-error methods by brealdog up the equation into three simultaneous equations, as follows:

+ U%(TB

- Tc)

+ -

(ta t*)CpR In equation (33), the term LHV.+ q~ a t 1 is the total sensible heat released to the furnace per pound of combustion products and may be replaced by qpA, the adiabatic sensible heat. q ~ may , be read from Fig. 2 at any assumed vdue of T g . With these simplifications, equation (33) reduces to ,-..A

(32)

The solution may then be achieved by assuming vdues of T B and plotting solutions for equations (35) and (36). The point of intersection of the two equations is the solution. T A may ~ be evaluated by cdculating the adiabatic sensible heat

The term U s- w ( T E - T c ) in aquation (32) is Sc generally negligible except for rear waIl impingement effects. It is convenient to drop the term at this point and correct for the effect later. For most marine boilers the temperature of the radiant heat absorbing surface (RHAs~is close to 1000 R (540 F ) , so the term Tc/1000 is approximately unity. Since the value of Tp/1000 is between 3 and 4, the relative value of ( T ~ / 1 0 0 0is) ~so much higher than 1 that the term Tc/1000 can be taken as equal to 1 with little error. Further noting that T P = (TA.TE)112,equation (32) becomes

then Tnf may be read from Fig. 2. Usually, the convective term in equation (25) is negligible; but when a rear wall ie fitted, especially in a shallow furnace, the convection effect of the flame blasting against the rear wdl may be significant. An equation for the surface heat transfer coefficient WRW, based on the actual surface exposed to the gas, is

Where = Prandtl number k GD - = Reynolds number F

THERMODYNAMICS AND HEAT ENGINEERING

MARINE ENGINEERING

invariably inside the tubes, with some type of extended surface outside. There is a large variety of extendedsurface types, ranging from cast iron fins shrunk on steel tubes, to stud fins, aluminum fins, and spiralwelded steel fins. Such elements are proprietary in nature and performancedata must be obtained from their

49

manufacturers. ExtendedeUrface perfomance data are usually acquired by tests of the particular geometric design. See Subsection 3.12 of Chapter 3 for an example calculation which illustrates the considerations involved in designing boiler heating surfaces.

Section 3 internal Thern~odyna~nicsof the Steam Turbine 3.1 Nozzle Flow. A nozzle is a short flow passage of converging or converging-diverging flow area whose function is to convert thermal or pressure energy into kinetic energy. It thus forms an essential feature of both steam turbines and gas turbines. As the fluid passes through a nozzle, no external work is performed, and no heat is transferred, so that the general energy equation reduces to

LOSSOFPRESSUREB~D~GRADAT~ON OF ENERGY AT INLET

Fig. 13 Temperature f ador

Table 1

Tube Bank Depth Correction Factor, FD

EXIT PRESSURE

upstream. This degrylation is evident as a loss of stagnation pressure and, hence, of the pressure difference available to cause flow through the nozzle. ~t is indicated on the enthalpy-entropy plot of the nozzle process shown by Fig. 15. There is degradation of energy within the nozzle itself, so that the exit velocity is not as high as ideally ENTROPY, s possible. The total degree of degradation is expressed Fig. 15 The nozzle flow procer by the nozzle efficiency, which is thus the ratio of the energy actually converted to kinetic energy to that theoretically possible. In equation form, the definition of the nozzle efficiency q N is such as,the angle through which the fluid is turned, nozzle dimensions, and the ratio of approach kinetic q N = C?/(&~J) (50) energy to the total kinetic energy developed. Empirical h00 - hl' curves, such as Fig. 16, give nozzle efficienciesfor bladeThe meaning of hl' and the derivation of equation (50) type turbine nozzles for dry or superheated steam. T~ are evident in Fig. 15. determine the efficiency of a nozzle (either fixed or An alternative designation of the degree of energy moving), the basic nozzle efficiency h-2 and height tor~ ~ ~ ~ e r is s i given o n by the velocity coefficient kN, rection factor f~ are read from Fig. 16 and the nozzle which is the ratio of nozzle exit velocity to that ideally efficiencyis computed as

A, calculate L from equation (47) and multiply L by

ST= tramverse pitch, inches

sL= longitudiial pitch, inches

pressure of CO2 and HIO- Enter pR,the total ~ i 14~at the . average gas bulk t e m p e r b e , and at the proper p R value, ~ read € Q on the left scale. Then reenter at the gas bulk temperatme and read h?/rQ On the right scale a t the appropriate tube surface temperatme. hr is equal to the product E Q X h r / r ~ iincluded in the result is a tube surface emissivity of 0.g5. the 2.4 Heat Recovery Equipment That portion hest transfer equipment that absorbs heat at ternperatures below the saturation temperature of the generated steam is considered to be heat recovery equipmenta the Generally, such equipment absorbs the hest

d = tube diameter, inches l-he e ~ s s i v i t y of flue gas is a function of its temperature, the mem radiating length L, and the pressure Pa of its rdiating constituents (pfiwatervapor and carbon dioxide). The flue gas fuel oil in 15 percent produced when burning p ~ t i a pressure l of 0.114 excess air has a water vapor atm/atmand a carbon dioxide partial pressure of 0.125 combustion air (gas air heater) or into the incoming atm/&tm, a total of 0.239 atm/atm. Values of r~ for feedwater (economiser). Combustion air heaters are generally of either the this mixture are plotted in Fig. 14 for a range of gks bulk temperatures and a range of P R Lvalues, where PRis in rotary regenerative type or are tubular, with air atm/atm' and L is in feet. The curves are usable from through tubes heated by combustion gM p a s a d mound the outside of the tubes in Cross flow- The mtar3' re10 to 20 pementexcess air without appreciable error. and performance data is generative type is Plotted on the same figure is the value manufacturer. Tubdm best from the proposed air heaters can be readily evaluated by the methods reviewed in the foregoing for tube banbe ~ ~ t ~ ~ d ~e~onomizers d - ~ ~ are ~ femployed a ~ e of equation (20). TO determbe to the exclusion of bare-tube units- The feedwater which is another

(51) '

It may be seen that k~ is simply the square root of qN. The nozzle velocity coefficient is a function of factors

kN2 = fLk2 (52) For wet steam, a correction is necessary to,account for the impingement of the slower moving droplets of water on the back of the blades. l-his correctionis taken by some authorities to be

1 111

MARINE ENGINEERING

"

'CHERMODYNAMICS AND HEAT ENGINEERNG NOZZLE PARTITIONS

-

NOZZLE HEIGHT

' *' ''

8 RADIAL CLEARANCE, IN. D-CLEARANCE DIAMETER, IN. C O N T R A C T lFACTOR ~~ DUE TO THROTTLING OVERALL LABYRINTH PRESSURE RATIO FACTOR

a

NOZZLE OPENING

Fig. 20

N o d e nomenclature

Fig. 19 Blade and nozzle partilion nomenclature far a typical converging

'k

nozzle, section taken at mean diameter

C1, = tangential component of steam velocity leaving and blades is a relative one, nozdes being considered as fked and blades (or buckets) as moving. Fok nozzles or blades below the critical pressure.ratio, the area at exit is of primary importance. From Figs. 19

in the plane of the turbine wheel. Let m =

nozzles d = diameter T denotes the tip of nozzle M denotes the diameter of nozzle R denotes the root of nozzle

- -- -C=l r-T W 8

As

A = L d ~ a m Esin al

..

(64)

,where A is the area at exit from the nozzle, in square inches; L is the blade height in inches, and d~ is the mean wheel diameter. al is the angle of the steam to the plane of the wheel. Usually there is a small difference between the actual steam angle, all and the geometric angle a!. This angle arl - a; = 8 is known tis the deviation angle and is a function,of both the angle through which the steam is turned and the Mach number, and approaches zero as the Mach number approaches 1.0. In equation (64), al should be used when its value is known.

VlnT

A C C E L E R A ~DUE ~ ~ TO ~ GRAVITY- 32.2 F T , / s ~ ~ >

PRESSURE BEFORE LABYRINTH, LBS/FT2 PRESSURE AFTER LABYRINTH, ~~~.~~2 STRIP THICKNESS ATTIP, IN.

.

PITCH OF STRIPS, IN.

CIZM= C z v l n ~

VlnR

'

yln 7 specific volume.of steam leaving nozzles A, = axial flow area in plane a t wheel

W~ =< total weight of steam flow

't

"

CARRY-OVERCORRECTION FACTOR FOR STRAIGHTTHROUGH LABYRINTH; UNITY FOR STAGGERED

,

.

The result of these requirements is a warped blade, with generally pure impulse at the root and with a large degree of reaction at the tip. Normally, the laat few stages of the LP turbine are based on the free vortex condition, with the other stages having reasonable approximations of this flow. Obviously, as the turbine size incremes, the blade lengths increase and the free vortex design may be extended into higher-pressure

MARINE ENGINEERING

54

THERMODYNAMICS AND HEAT ENGINEERING

q-he rnmbined windage and friction losses may be estimated by Kerr's equation as follow:

A h

z

hoO- h*'

I,

--

(72)

1 f hllo pressures Po for each stage are known, a they

bo in the design process, then a stage-by-stage I ) I ~ ~ of ( J the conditions for each stage can be made on the Mllllitrr chart (h-8 plane; see Fig. 18). When completed, 11ll)t is known the state line, or condition line, for 111. bllrbine, shown by Fig. 22. It is I ' * I ~ I ~ in the andysia of extraction point conditions, "L'l"r requirement is knowledge of enthdpiea @A flitlotions of Pre8sures. However, the end of this f1111fl(r

d VI

w

hp = horsepower 10% d M = mean blade ring diameter, ind a = b k diameter to root of blades, inE = peripheral admission fnx?tion L = blade height, in. iy = blade speed, fps

55

h e , known the state h e end point (SLEP), represents the static enthalpy only, whereas the stagnation e n t h d ~of~the exhausting steam must be known for e such Uses 88 condenser design. Thus the C -- component 2gJ is u s u d ~ added, and the resulting stagnation enthalpy at exhaust is also plotted on the state line diagram at exhaust Presswe. This component is that which "presents the approach-velocity h p u t to a followkg stage, in equatioq (62), but which, in the last stage, must be wasted. ~t thus forms the major part of the t u b h e leaving or loss.

Sectio~~ 4 External Thenodynanrics of the Steam Tflrbine ''I

IWe Line for the *@' a hrbine unnll the

a

unit. The wheel horse- power output, thmttle pressure and temperature,and is given by equation (68). exhaust pressure are sufficientfor this task. of Wa, WL, hoop hsO,and hpj A preliminary step is to express steamflow in unitfom @@Qmowllt. The total Power delivered into the turbine a a steam rate (or waterrate) thusly: @h@fb i h must ~ be expressed zw a summation of the

compounded stages

who01 horsepowers by

ENTROPY

fig.22

where

propulsi~nturbine state line

SR

=

W,SHP

2544 UEtW= SRVM

UEw = heel Used energy (see Fig. 22), Btu/lb 9~ leakage efficiency

- hr4 - hpf

(68)

= total steam flow entering stage, lb/hr

wL = leakage

flow, i.e., flow that byk'asses the or moving blades, lb/hr hp = power absorbed by windage The wheel work per pound, based on total flow, is

enthalpies. stage efficiencyis then the ratio of ~h~ work delivered to the shaft to the available stage, or

, =

or, as a close approximation

per

(74)

(75)

MARINE ENGINEERING

59

be read from the intersections of the shifted pressure lines with the nonextraction state line. The approximation outlined here is not quite adequate for reduced-power conditions, or at unusually large extraction flows. Other techniques, such as discussed next, must be used. 4.4 Lambda Ratio. For large variations in flow such as occur when reducing to 80 percent power Or less, it is necessary to account for the change in efficiency because 'peed of the change in the ratio of blade speed to in SucCeS~ivestages. This can, of murse~be done by returning to the original design and applying the theory in Section 3 again. However, prodiscussed for a computer this is a tedious operation and normally carnot be accomplished in a timely manner, especially for preliminary work. The designer must (gq) therefore resort to other techniques based on external

that most desigoers and turbine builders prefer to use t6e s t r ~ g h tLine. since the state line is ~ m primarily d for heat balance work, an error of 6 ~ t inuestimating the e m in enthdpy at a given point resultsin a extraction flow. T~ arriveat a satisfmtory state line for fdl power, it is then only neoeasary to comect, on a Mollier diagram, the point of idtial PreBme and enthdpy, ho, with the enthdpy a t 90 per cent of throttle point of pressme, and the state line end point (see Fig. 22). ntraction ent.alpies csn then be read a t the appropriate shell or stage pressure. The steamleaves the tmbine at a total enh, = hi

+ EL + (RL)(SHP)(2544)

I

since the pressure a t the condenser is very low, the

Every turbine stage has a value of U/Cf for which its

preame ratio from the point of interest to the condenser efficiency is a maximum (this is shown by Fig- 18 of is typically supercritical, and maximum flow exists Chapter 5). For example, it is 0.5 for an ideal impulse

for the pressure a t that point. Under such conditions, stage. And although a propulsion turbine Consists of a flow theory predicts that the flow parameter number of stages for each of which the ratio may be W 6is and in f m t hm a value of approxi- different, as when impulse and reaction stages are used in the same machine,' there is always some vdue of mately 0.40 for superheated steam when To is in degrees d ~ Nfor whi& the efficiency of the entire turbine R, po in psi, W is in Ib/sec, and A is in squme inches( ~L ~ Z~ ~ 1 Further, for modest changes in conditions at a point in is a maximum (Z implies sumation Over all compared to the the the in is his parameter is known as Lambda, and is convenchange in Po, and A is fked, SO that the relation r pofouom. m e premure a t a point should thus tiO*lly expressed to the flow from that point to the be pediction is found to be essentially condenser, and true in practice; it is further theinpressure is wherein the constants, kcluding 0.5 for U/C', are pmportiond to the flow pastfound the that point question. included in the numerical mefficient- The efficiency at This additional distinction is necessary because some of function of the Lambda a point may be extracted downstream, ofi-design points is a ateam ratio X/Xo, i.e. and therefore does not reach the condenser. The principle stated in the foregoing is used to find shell pessures a t extraction points, and from them the extraction enthalpies following small changes in flowThe state line does not shift significantly because of Fig. 26 is a plot of the relative efficiency ofofa impdse rsaeonable ortraction flows, 80 that the enthalpies can s g h t h/lo- This plot is

THERMODYNAMICS AND HEAT ENGINEERING

61

MARINE ENGINEERING

ICS AND HEAT ENGINEERING

BOILER 88% EFF

000 SHP ABS MAXIMUM

Eb ~ 0 . 8 5 2 7(SEE FIG.23)

a = FLOW, L W H R h = ENTHALPY, BTUILB. Fig. 34

Simple steam cycle

EXHAUST ANNULUS AREA = 25 FTP

f r ' I.O125(SEE FIG.24)

A € = 1481.2-907.5 = 573.7

R L = 0 . 3 5 ~1 . 5 ~ 0 . 5 LEAKAGEaMECHANICAL LOSSES = 3.5% ITERATION, STEAM FLOW = W PbxA

I

1163 .56 ~ 0~02 5=.4360 ~

BYadn per hour, divided

1

-

(I.o¶)(-)

= 163,600

Le/HR

**. E L 5 10.8 (SEE FIG.25)

SECOND ITERATION, STEAM FLOW = STEAM RATE =

E x s ~,Xf,xAEn0.8527x 1.0l.25 x 573.7-495.3 h,- .,E, 1481.2 495,3 = 85.9

hi

(I .04) (245i,:

!:60p)=

[63,800 LBIHR RL X 2 5 4 4 hc. hi + EZ+ STEAM RATE = 999'0 Ag. 35 S t a h one and steam rate for cycle cafculationr

~

~ =, 5.46 ' LB/SHP-HR ~ ~ ~

the m~chanicdequivalent

t# b a d and the efficiency. Thus

divided by the net or 15,180~30,000= O e 5 0 6 ' lb/sh~-hr. The heat rate is the quantity of *heat t o produce one horsepower per hour and 144APvfQ~ t u / h r ~ ~ ~ u l aby t edividing d the net heat added to the plant, Per hour, by the horsepower produced.

MARINE ENGINEERING

'I'llo quantity of fuel required is determined by dividing l d I ~fr~el ~ j heat output by the boiler efficiency of 0.88 and

IIIIII higher heating vdue of standard fuel of 18,500 llbl~/lb,to which is added 46 Btu/lb to account for the ti~rlnibloheat added by the fuel oil heaters (100 deg F rim tdt 0.46 specific heat) :

Ipuel required = 2509846,193 = 15,370 ,b/hr (0.88)(18,546) I)ividing by the 30,000 shp output, the specific fuel ~~t~ll~nrnption is found to be 0.512 lb/shp-hr. Sa2 The Regenerative Cycle. The power cycle shown 111 Itig. 36 is complete, but certain problems would arise If oh a cycle were used. The feed temperature is unkr('melylow; a result, the economizer in the boiler wclllltl condense sulfur ~roductsfrom the flue gas, which wnrlld cause corrosion. Further, the feedwater would IIELVI! high oxygen content (no deaeration is provided), wal,trr wlriuh ~ides. would cause corrosion and pitting in the boiler

1

200 F Q = 1050

Q=

---~ g 36 .

FLOW. LB/HR

h = ENTHALPY, BTU/LB

Simple steam cycle with parasitic loads

ta&. ~ ~ ~ j + b l ~ - ~ and system leakage are bssed on reference [gl. soot,blowing requires 760 lb/hr, and system l e h g e losses are taken as 1/2 percent flow, or about 900 lb/hr. Both of of the mustbe replaced by makeup these items lost feedwater introduced to the condenser. ~h~ t h e r d enerw added to each pound of water by the feed pump is the same in the example, so that the thermal energy added to 177,920 lb/h. is 1,060,384 Btu/hr.

Uowance must be made for Pressnue and temperature drops in the main steam line. It is customm to allow up to the nearest about 2.5 percent on Pressure, 5 psi, and 5 deg I? for temperature. Thus, the heater outkt conditions are taken to be 875 psig and 955 F. NO dlowance for loss is made in the deSWerheated system, so the desuperheater outlet enthalpy is is 1250 Btu/lb- The total is 2300 lbhr' 175,620 lb/hr and the demperheated The flow of heat in the system illustrated by fig. 36 may be tabulated follows:

1l)rltrainedoxygen and air can be released by bringing tlrn foodwater to a boil. By using steam, bled from the t\~rl~iaas, the feedwater can be raised to the boiling I@lll~lorature and held there in a deaerating feedwater I r e ~ b in r an efficient manner as the bled steam has al~entlydone useful work in the high-pressure turbine befo~~o being used for feed heating. Illtrod feed heating may also be done after the dercrr~~l~ing feed heater and feed pump. For highest rflinioncy, there should be N-1 feed heaters, where N (r lllro number of turbine stages, since this leads to lncxirnum regeneration, but such an array of heaters &r&dbleed points is not justified in marine service. flbltbionary practice employs an extensive number of Iie&llrrr8,but such plants are not restricted by the space llmill~~tions of a ship's engine spaces, and they develop

generally used in cycles where the feedwater is heated to a maximum of about 285 F, so that economizers may be used for heat recovery in the boilem. Where feed temperatures are higher, gas air heaters ape used. Consider a simple single-heater cycle, using a deaerating feed heater, and otherwise identical with the cycle of Fig. 36. Steam could be bled at the crossover pipe between the H P and LP turbines at 60 psis and 1243 Btu/lb, and led through a pressure set at 46 psia, to a deaerating heater. The bled steam and the incoming feedwater could be sprayed together resulting in a saturation temperature of 276 F, 245 Btu/lb enthdpy. Bleeding steam from the crossover would reduce the horsepower developed since less passes through the unit. To compensate for this, the throttle flow must be increased. Let QI = the quantity of steam bled in lb/hr A&, be the increase in main throttle flow required. m e n , using the figures developed in Fig. 35, the reducis tion in heat available to the turbine lost heat = Qr(1243 - 996.7) and the flow needed to replace this heat is AQt(l481.2

AQt

.............a.

.................. ~ o t a l................................ .

175,620 X 1483.5 = 260,53%270Btu/hr 2,300 x 1260 2,875,000 263,407,270 B t u b 177,920 1 b b

r:

0.5084 QI

Then, leaving the exhaust of the LP turbine, the steam flow would be 163,550

+ AQt - QI

=

163,550 - 0.4916 QI.

Employing the procedure used in Section 5.1, the heat entering the gland exhaust condenser is

I b b

Btu/lb

Btu/hr

Leaving the main condenser 176,325 - 0.4916Ql 58.7 10,350,278 - 28.8691 Air ejector intercondenser. . 245 1250 - 93) 283,465 Air ejector after condenser. 245 1250 - 168 265,090 Gland exhaust condenser. ., 300 (1281 - 1681 333,900 Total.. .............. 176,325 - 0.4916Q1 11,232,733 - 28.8691

.

Fuel heating at 200 F . . 1,050 ~ b d dram a t 200 F................ 300 Air ejector after condenser drain a t 200 F....... 245 ~ ~leaving t dsurge t and entering feed pump8 177,920 Feed -p ........................... 177,920 .fotd to boiler.. ......................... 177,920

- 996.7)

Equating these gives

I

The boiler output is t b n .

65

THERMODYNAMICS AND HEAT ENGINEERING

--

THERMODYNAMICS AND HEAT ENGINEERING

MARINE ENGINEERING Ib/hr Btu/lb Leaving gland exhaust ,denser. ............. 176,325 - 0*4916&1 168 After condenser drain ..... 245 ~ h condenser d drain .... 300 168 +&I 1243 Bleed flow.. .............. F.0.hater drain. ........ 1050 168 ~ ~ t r r .............. l.. 177,820 4- 0.50&1Q1

Since the DFT enthalpy is 245 Btu/lb, (177,920 + 0.5084~~)245 = 11,500,693 + 1214.14~1 QI = 29451 lb/hr The total flow to the boiler is

+ 6) = 48,416,143 B t u / b

to the boiler, and the boiler output is (192,893 - 2300) (1483.5) (2300) (1250) Total boiler output less feed input Net heat input to boiler

= 282,744,716 Btuihr

2,875,000 = 285,619,716 48,416,143 = 237,203,573 Btu/hr

After condenser.. ....... 245 G h d condenser drain. .. 300 Bleed flow. .....-...... NXo fie1 oil heater drain.. ... DFT outlet flow. ....... 177,560

+ lW1 11,500,693+ 1214.1+@1

1.04

+

=

41,:1!J0,787 Btu/hr.

The boiler output is ,

(192,537 - 230011483.5 = 282,216,590 ~ t u / h r (2300)(1250) = 2,875,000 Total output = 285,091,590 ~ t ~ / h r input = 48,326,787 Net heat from boiler 236,764,803 ~ t u / h ~

firbogenerator Throttle steam conditions.. .850 pig, 950 F Exhaust. .................

13,963

and the horsepower developed by the HP turbine

+

-1- (1481.2 - 1243)(149,322 29,451)/2544 = 16jog5 1.04 and the total is 30,058 shp~ hthe reduced ~ ~ steam , flow in the LP turbine slightly increases the turbine efficiency, and the calculation be repeated with a new ratio of AQr/Qz and a new nonbleed flow. Since the exhaust enthalpy has been changed, the equivaJent nonbleed water rate is :

=

Dividing the net boiler heat input by a fuel heating value of 18,546 Btu/lb and a boiler efficiency of 88 gi~a fuel oil requirements of 14,5341b/hr. Dividing by the 30,000 shp o ~ t ~ u ~ r e s uinl tas specific fuel rate of 0.4845 lb/shp-hr. Tbis is a saving of 5.4 percent over the simple cycle, and in addition the boiler is protected from corrosion. A further gain in efficiencyaccrues in this cycle. In the high-pressure turbine, more steam Passes than in the nonbleed condition, and less passes through the lowannulus is the same prerrnve tw~lI1e.Since the before, the volume flow is m-hced, and the h3aving velocity be less also. The apparent exhaust flow is l63,30 - 0.4916 x 2 ~ 5 = 1 149,322 lb/hr (apparentflow is t h t t l e flow less any bleed but including the gland leakoff steam). The exhaust annul- is

(177~600+ 0.5095Q~)245= 11,4799561

176,400

1 (1243 - 995.6)(149,322)/2544 -

Assuming the same pump efficiency, the feed P U P per pound of wateris unchanged (iVe.,6 Btu/lb), so that the generator load should have increased somewhat, but this can be balanced by the decrease in boiler forced-draft blower power requirementsThe total enthalpy of the feedwater is 192,893(245

11,232,733 - 28.86Q1 41,160 50,400

25 sq ft, SO 149,322/(1.5)(25) = 3982, and from Fig- 25 the exhaust loss = 9.7 B ~ and ~ hw/ = g85*g ~ ~+ 9.7 = 995.6 Btu/lb, vice the 996.7 Btu/lb &own by Fig. 35. Then, the horsepower developed by the LP turbine is

177,920 + (0.5084)(29,451) = 192,893 1b/hr

Fig. 37. Figure 35 outlines the calculations for the propulsion turbine. This calculation, and those sum1214.21Q~ marked below for units,use the metho& and data

'I'I"' DPT outlet e n t h a l ~ is ~245 Btu/lb; therefore

Btu/hr

Iso4

W R = 1481.2

2544 = 5.448 Ib/shp-hr

- 995.6

Then, the throttle flow is 5.448 x 30,000 AQt = 163,440 + AQt and A Q ~= (247.4/485.6)&1 = 0.5095Q1

+

and the exhaust flow is 163,440 - 250

+ AQ, - QI

=

163,190

0.4905Qr

where the 250 lb/hr is the gland leakoff steam. Leaving the condeqser: Main turbine exhaust. 163,190 - 0'4905Qz Turbogenerator exhaust. Makeup feed. ................. lS6O Air ejector drain. ............... ,175,965 245 - 0.4905~1lb/hr ~ ~ t ............... d.. ......a

....a*.

'I'll~~ preparation of a heat balance is usually the first ?*P in initiating the design of a steam propulsion plant. results of the prelimioary heat balance are the ki~llmentalinput to purchase inquiries, and also for lllfill Plant desifP tasks rui sizing of piping. The first !wJ balance may be done from the approximate data in ~ ~ f c m n c[gl, e but subsequently, data supplied by the @f)tllponent vendors is used to update the calculation. In the last example in 5.2 a direct solution of the heat balhaoe problem was presented. Obviously this problem Would be more difficult if several bleed points are needed nlwl rrlore heaters employed, especially if the bleed presU l l t ~vary with flow. The problem becomes even more @~l~ll~lOx if ships's service steam is added to the balance. Wlliln a direct solution of the heat balance is possible by b i @ i l la~ series of simultaneous eq~ati008,it i s generally Illor0 ~impleto use an indirect trial-and-error solution b,Y na~uminga condensate flow leaving the main con-

rated capacity.. ........ .........13,600 lb/hr 1135 Btu/lb Exhaust enthalpy.. consumption at 480 gpm and 1200 psig. . 12700 lb/hr Exhaust Bnthalpy at 480 gpm and 1200 psig.. . .1139 Btu/lb Main air ejector steam consumption 1st stage.. ............. .245 Ib/hr 2nd stage. ............. .245 lb/hr Intercondenser drain temperature.. .......... .I25 F After condenser drain temperature.. .......... .200 F . Steam supply at 150 psia. . .I250 Btu/Ib ~ i ~ t i plant l l i ~ Water production.. ....... .11,400 gpd steamconsump~onfrom low-pressure bleed at

Setting up in tabular form:

Total makeup feed.

+ O.M)gWr

-

...

.3330 lb/br

consumption) ...........lo00 l b / b

1

69

MARINE ENGINEERING

-------

TURBO-OENERATOR LOAD

L, , ----- -

The f i s t step is to estimate optimum bleed points. The feed temperature leadng the deaerating feed tank has been set at 280 F to prevent the condensation of sulfur products from the flue ges in the boiler economisers. To achieve this temperature, a pressure of 49 p$a must be available from the auxiliary exhauatlintermediate pressure bleed system. This pressure is controlled by a pressure regulator installed in the bleed systems. To provide heating steam when no bleed steam is available (as for example, when going astern), makeup steam is supplied from the desuperheated system through a pressure regulator set at 45 psia. At certain times, there may be too much auxiliary exhaust regulator, re set at 53 psia, will steam, so a b a ~ k - ~ r e s s ~ - dischmge excess steam to the main condenser. Thus the auxiliary exhaust system can fluctuate only between 53 psis and 45 psis. This limit should be sufficient to prevent the feed suction water from the DF'T from flashing during maneuvering, especially if the DFT is placed well above the pump (40 to 75 ft). Allowing a 7 percent pressure loss through the bleed/exhaust system, the bleed steam at the turbine must be at least 7 percent hi&er than the desired 49 psia, or 53 psia. operation to be It is further desirable for this able to continue bleeding to at least 65 percent flow in

the main turbine. Since the bleed pressure at any stage is dependent on the flow through that stage, it is desirable to select a bleed point at (53/0.65) = 81 psia. Note that since this is a direcbcontact heater, there is no terminal temperature The optimum bleed difference. point for the low-pressure stage is then selected so that the temperature rise in the condensate is evenly distributed between heaters. ~ e a v i n ~ the condenser a t 1.5 in. Hg abs pressure and 90.7 F, the condensate will be heated by the air ejector intercondenser and after condenser and the dand condenser to a temperature of about 100 F. The temperature rise to the DFT is 180 deg F, approximately half of which should be achieved in the LP feed heater. ~ h u s ,a proper condensate temperature leaving the LP feed heater is 190 F. Since a 10 deg F' terminal difference is usually needed between the heating steam and heated water in shell-and-tube-type heaters, the steam entering the heater must have a saturation temperature of 200 F, corresponding to 11.5 psis at the heater shell. With a 10 percent pressure drop in the piping, the turbine bleed point pressure must be 12.8 ~ s i a . For the preliminary heat balance then, the LP bleed point pressure is 12.8 psia at an enthalpy of 1138 ~ t u / l b and the I P bleed point pressure is 81 psia at an enthalpy

lbb

Btu/lb

Btub

154,865

58.7

9,090,576

283,465

265,090

154,865 lb/hr

Note that in this calculation, the weight of the drain is not added to the total, since the intercondenser &&n goes to the condenser and the after condenser drain to the freshwater drain collecting tank shown on the

THERMODYNAMICS AND HEAT ENGINEERING

MARINE ENGINEERING

Entering the system lb/hr Chde-te from LP htr.. .......... 154,865 Drains from drain tank. ............ 30,175 Exhaust from feed pump. .......... 12,700 Feed pump recirculation............ ~lOO0 Bleed steam from turbine. .......... Q Total. ...................... 217,740 Qa system .................. (217,740 + Qe)

+

(217,740 4- Qc)249 = 49,471,470 Q, = 4,650 lb/hr 14340 X 16.5 X 0.2445(278 1270 - 277 = 10,370 lb/hr

+ 1270Qc

- 100)

Btu/lb

Btu/hr

158 24,468,670 180 5,431,150 14,465,300 255.3 5,106,000

1139

1,270 Qa 49,471,470 + 1,270Qe (217,740 + Qc)249

1270 249

'"-

-

(ho h ~ = ) 17,030(1481.2 1138) 2544 2544 = 2,297 whp and the wheel horsepower of the IP bleed flow is:

= 1,247 whp Total = 31,242 whp

EnteFing the subsydm

lb/hr Condensste from mitt air ejedor after condenser.. .................. 154,865 300 G h d steam from turbines. .......... 245 M . e air ejector after condenser drain. . fistder mr ejector dram. ............ Makeup feed, taken at 75 F . ......... 3,330 Air heater drain.. ................... 10,370 Miec. heating drains.. 1,100 Low-pressure feed hater.

Btu/lb

1281 168

9,639,130 384,300 41,160

............... ............

+

h v i the ~ subs &em From gain t a d at 212 F............ (15,595 QL) 180 2,807,1°0 From LP heater at 190 F . 154,865 158 24,468,670 Total. ..............................................27,275,770

...........

+ I80 QL + 180QL

Equating incoming to outgoing total enthalpies gives the air heater drain in the air heater coils. It should also be noted that, in many cases, the low-pressure feed heater is drained, via a drain cooler, to the main condenser. Also, sometimes, the entire drain tank is also drained via a drain cooler to the main condenser. The steam supply to the steam air heater is bled from the 81 psia stage at 1270 Btu/lb. Allowing 7 percent pressure drop in the piping, the pressure at the air heater is 75 psia when rounded off. At 75 psia the saturation temperature of the steam is 308 F and the condensed drain enthalpy is 277 Btu/lb. There must be a temperature difference between heating steam and heated air leaving the air heater; this terminal difference should be between 25 and 35 deg below the steam temperature. Choose, for example, 30 deg F aa a terminal difference,so

+ 1138Q~ = 27,275,770 + 180&~ QL = 14,580 lb/hr Then, the drain tank flow is 15,595 + 14,580 = 30,175 lb/hr at 212 Fj and the condensate flow is 154,865 13,307,070

lb/hr. The next part of the system is the deaerating feed heater (or DFT). This unit receives condensate from the LP heater and drains from the drain tank via the drain transfer pump. I t also receives heating steam from the feed pump turbine exhaust and bleed steam he bleeder from the intermediate pressure bleed. steam is controlled by a pressure regulator set to maintain 280 F. A weight flow and heat flow balance for the DF'T gives

-1139)(12,7m) = 6.3 Btu/lb, 222,390

-

W

145 045

-=-L=

P6.A

3868

1.5.25

= 27,698 whp

is made for liquid compression. The fuel rate is now calculated from the heating value

18,546 \

n@ wll001 horsepower of the LP bleed flow is:

+ L16.5 X 0.2445 X (278-loo)] =

19,264 Btu/lb of fuel

The heat added to the superheated steam is

MARINE ENGINEERING

185,520(1483.5

'THERMODYNAMICS AND HEAT ENGINEERING

- 255.3) = 227,855,660 Btu/hr

1 Ill1 rl~l1t111,ted steam,

II I ' r

The enthalpy of the desuperheated steam is 1250 Btu/lb, and the heat added to it is

and a low-pressure turbine. The

1 I' unit8 are mounted on the same shaft in the

tlg, with their high-temperature ends back to 'I'h~sthere are only two input pinions, as with Ill0 crc,l~vc!ntional nonreheat turbine. I1rflflit:Cionsof a reheat turbine state line during E b = 0.865 lflulll1lillllJ'.Y design studies are more difficult than for llod the temperature correction may be computed aa l l ~ ~ ~ l t ~ ~ Iturbines, 1 ~ ~ 1 1 ~ 0 since there is a wider range of 17l1tl 1 nt\(,crri~tics that affect the line. The marine engineer T 4100 ft = r l l l l ( tolll! burbine designer must cooperate more closely to ~ l l ~ l l l l ~ c ! e11r design acceptable to both than is necessary 5000 (96) f t = 1.01 11 nil 11l)lor propulsion plants. Nonetheless, as a first R a l l ~ ~ ~ ~ of l ( la ( ' suitable state line, the high-pressure turbine The state line energy for the low-pressure turbine is ~ ; ~ l ~ r r1)rossure ~ ~ a b can be selected at about 20 percent llr I 141~rottle pressure, and a state line can be conE E L= EbftAE (97) l l ~ c l t ~ f l twith l an efficiency of 70 percent (excluding E B L= 449 Btu/lb ll~fll.11lg lblld gear losses). A 10 percent loss of pressure With an astern turbine loss of 0.5 percent [see equation i l l ( 110 rfrI~oater may be assumed. The balance of the (78)1, a first estimate of the steam flow is determined to be 138,400 Ib/hr from equation (83); therefore c l l ~ l l lc ~ !tl~i

Ijlllll{,

16,870(1250 - 255.3) = 16,780,590 Btu/hr For a boiler fitted with a steam air heater, a boiler efficiency of 88.5 percent can be expected; therefore, the fuel burned per hour is determined to be 227,8555660 16,780,590 = 14,349 ib/hr (19,264)(0.885) and the fuel rate is 14,349/30,000 = 0.478 Ib/shp-hr. The results of the foregoing calculations are entered on the heat balance diagram shown in Fig. 37. Since the shp check was close and the f i s t estimates of the steam consumption by the feed pump turbine and the air heater were well confirmed, these results can be considered final. If any of the checks had failed, the process would be repeated with revised estimates based on these results.

73

+

ENTROPY

RU. 38

&timote of state line for propulsion turbine with reheat

Section 6 Waste Heat from Diesel and Qas Turbine Engines I

from Diesel Waste Heat. A large fraction illput to an internal combustion engine is r @ a l e ls* ~l!ll*ibleand latent heat in the exhaust gases. *h+rrllpr l l u 1 ~fii~llificantfractions are lost via cooling of )fi@lvl, nlLtcr1 lube oil, and inlet air (turbocharged nlllle@ "'ill1 ~bftercoolers only). For example, the f ~ ~ ~ ~Ilt!atin~ut l l ~ l going l ~ ' into ~ ~the waste heat streams hf H k\~l'boal~larged two-stroke engine might be (1

#f tl!fl

Ilfldl

0.35 to exhaust 0.15 to jacket water 0.05 to lube oil 0.05 to aftercooler I I!r ~llnnt!loxlraust gas temperature is a t least 600 F

*&11111 111bf1,

it is feasible to extract part of its sensible ~lro(lllc~ usable steam. As the cooling water I E + f l l l ~ l ' i l l J l l ' ( i~ ! 1 ~than ~ s200 F, there are few uses for this *' I ' One use of practical importance, how-

ever, is the operation of a vacuum distillb for freshwater production. Steam can be produced in a heat exchanger (waste heat boiler) in the exhaust duct. ~h~ maximum steam pressure 9 b t ~ n a b l eis limited of course by the exhaust gas temperature, but othelqrise the premure is set by considerations regarding the use of the steamand the quantity needed. If steam is to be used solely for heating purposes, a relatively low pressure, say 15 psig, may be adequate, but usually the heat available is far in excess of lowpressure heating needs. Often, the ship service electrical needs at sea can be met by waste-heat steam applied to a turbine-generator. The higher the steam pressure, the lower will be the turbine steam consumption, but also the lower will be the quantity of steam that can be produced. Figure 39 illustrates alternative steam production at 50, 100, and 150 psig, showing that with

MARINE ENGINEERING

THERMODYNAMICS AND HEAT ENGINEERING

MINIMUM TEMPERATURE DIFFERENCE("PINCH POINT*).

o

b~ 4s

EXHAUST GAS FLOW 119.000 LBlHR

Z 4 6 8 1 0 1 BACK PRESSURE, INCHES H,O

2

IRect of back preoure and intercooler outlet tcwnperctture on exhaurt tmpbroturq Sulser RD-type engine

taken by the external cooling devices. The evaporator must not change the temperature of'return cooling water from its specified range under any condition of operation. 6.3 Use of Gas Turbine Waste Heat. Exhaust gas heat from gas turbines can produce steam in the same manner as for diesel engines, and for the same purposes. Since g&8turbines are generally less efficient than diesels, the heat available tends to be greater than with diesels. In fact, there is sufficient energy available to suggest use of the steam in a propulsion steam turbine geared to the propulsion shaft in parallel with the gas turbine. Perhaps 20 to 35 percent of the total power can be produced by the steam turbine, with a consequent major improvement in the fuel rate obtained with the total system. The design objective in a combined gas turbine and

0.25 0.50 0 75 FRACTIONAL LOAD

pa. 40 Exhaust ROW and temperature, Sulzer RD engine

the same inlet temperatures more steam is produced at progressively lower pressures. The minimum temperature difference, or "pinch point," as indicated, is the governing consideration in the steam quantity that can be produced. However, additional &earn is sometimes obtained at a lower pressure in a second boiler downstream of the fist. The minimum temperature to which the exhaust gas 2 cooled is also a limitation, since the temperature should not be allowed to drop below the dew point in order to avoid corrosion in the cold end of the boiler. Wade heat steam systems are designed in a variety of forms, but generally contain the components expected in a self-contained system. The designer, in making a heat balance, will apply the same techniques outlined earlier in this chapter. He must allow for the fact that

0

ZOO

400

wo

800

EXHAUSTGASTEMPERATURE,.F

1000

76

THERMODYNAMICS AND HEAT ENGINEERING

MARINE ENGINEERING

1 D. Q. Kern, Process Heat Transfer, McGraw-Hill Iloolc Co., Inc., New York, 1950. fi W. H. McAdams, Heat Transmission, McGrawllill Book Co., Inc., New York, 1942.

LEGEND

=

- ---------

AIRORGAS SUPERHEATED STEAM LOW PRESSURE STEAM FEED AND CONDENSATE DRAIN , GLAND LEAKOFF AND VENTW STOP V. NCHECK V. 4 : ORIFICE ~

--

--

&coNTRoLv. P-OR-T

A BACK PRESS.V.

PG=PSIG PA= PSIA P = LBIHR FLOW h = BTUILB F = TEMP., DEG. FAHRENHEIT W = GAS FLOW, LBlHR

STEAM AND FEED CONDITIONS SUPERHEATER OUTLET 2 8 5 PSlG MAlN TURBINE THROTTLE 2 8 0 PSlG MAlN CONDENSER VACUUM AT 108.7.F FEED WATER TEMP. TO BOILER CALCULATED FUEL RATE

617.F 612.F I,NJ. 27.5 HG 260.F

(I "Standards for Steam Surface Condensers," Heat Il)xcitiungeInstitute. 'I "Standards of the Tubular Exchanger Manufarilurcrs Association," Tubular Manufacturers Associa-

,399 LBISHP HR

BASED ON

lioll.

MAlN TURBINE NON-EXTR ST. RATE 0.BLBISHP HR BOILER EFFICIENCY 46% HHV OF STANDARD FUEL OIL 19,650 BTUILB

H A. Egli, "The Leakage of Steam through Labyrinth

PIRJLIH," Tram. ASME, i935. FUEL F W , = 10.700

LOST 9-47

74I0F

(1 "Recommended Practices for Preparing Marine Htnrcm Power Plant Heat Balances," Technical and #e~tinrchPublication No. 3-11, SNAME. 10 A. Norris, "Developments in Waste Heat Systems klr Motor Tankers," Trans. Institute of Marine En&lrlnatwj, 1964. I I R. M. Marwmd and C. A. Bassilab, T h e l'lirrtnodynarnic Design of a Combined Steam and Gas

T'lrtdna fl7-[IT-16, 1967.

VACUUM PUMP

(5-

MAlN CONDENSER 2 . 5 " ~ABS. ~

MAlN FEED DEAERATING

Q =5 8 9 7 6

MAIN

CONDENSING PUMP

Fig. 45

Design-point,heat-balancediagram for a combined gas turbine and steam turbine cycle

steam turbine is fired by the exhaust gas. Observe also that the gas leaves the boiler a t 440 F, and thus still has considerable thermal energy available for the production of additional steam at a lower pressure. A second, low-pressure boiler is provided to make steam for the deaerating feed heater. The heat balance shown is for design power. It is also of interest to see how the important parameters change as the load is reduced. Figure 46, also from reference [ll], illustrates this. Actually, the effect on the system parameters is influenced by the manner in which the plant load is controlled. For the example given, the fuel flow to the gas turbine is controlled to

77

maintain a governed gas-generator rpm. The steam turbine is uncontrolled, with the output being determined solely by the energy available in the gas turbino exhaust. References

1 Thermodynamic and Transport Properties of Steam, ASME, 1967. 2 Joseph H. Keenan and Joseph Kaye, Gas Tables, John Wiley & Sons, Inc., New York, 1948. 3 Frank 0. Ellenwood and Charles 0. Mackey, Thermodynamic Charts, John Wiley & Sons, Inc., New York, 1944.

ASME

3001 20

-Fig. 46

I

40

00 6 0 \" 100 PERCENT OF PLANT RATING

120

hdanannof a ,..rhed gas turKne and steamNrKne cycle at fractional power

BOILERS AND COMBUSTION

C H A P T E R Ill

Sedion 1 Classifisation of Marine Steam Generators

tained during the record-breaking runs at about 30 ~ s i g , which was about the upper limit of pressure during the Civil War era.

1

Numbers in brackets designate References at end of chapter.

1 BOILERS AND COMBUSTlON

MARINE ENGINEERING

type of boiler. ~t is with this type that attention is focused primarily since it has been used most frequently since World War 11. while there have been many variations of the foregoing boiler types employed throughout the world, the typesdiscussed are fairly representative and provide an steam adequate backgmund for an understanding generator types and characteristics. 1.2 current lypes of Mer&."+ and Naval Boilers. the Past100 years steam pressures and temperatures have increased fmm 30 psig saturated to 870 psig-950 F in mostmerchant vessels, and 1200 psig-lOOO F maximum (950 F in mostpostWorld war 11 naval vessels. A trend is apparentin large, high-power installations where steamat 1500 psig-g50 F, and in some instances reheat to 950 F, appears desirable. ~h~~~ installations will be used in increasing numbers where economically feasible. For the mostpart, widespread use of water cooling in the furnaces is employed to reduce refractory mainheaters singly, or in tenance. ~~~~~~i~~~~and are used to obtain the desired overall steamgenerator efficiency. Attemperators are employed in most new construction to control the steam temper* operating range and thereby improve ture over a turbine performance. Desuperheaters are installed to provide low-temperature steam for audiary purposes throughout the ship. ~~~k~~ c residual oil is the most widely used fuel,

although in some instances diesel or other light fuel ofis are used. Steam-atomizing oil burners, first used aboard ships in the late 18001s,have returned to favor with the advent of high-capacity low-cost evaporators to supply the necessary water' This type of atomizer' while providing an extremely wide range of operation, results in a 1088 Of distfiled water which was, until recently, too big a penalty to pay for its advantages' However, improvements to reduce the consumption Of steam, coupled with abundant distil1ed water, have led to its widespread use, particularly in automated boilers. Two-drum integral-fumac0 a Two-Drum boilers, or D - ~ Y boilers P~ ss they are Often Of many and called1 are made steam drum and water drum connected by water and boiler bank tubes. Superheaters are instal1ed between the water screen and the boiler bank and may have tubes arranged either vertically Or horizontallyr depending in part on which arrangement best fits machinery arrangement- Where required, the temperature may be controlled by means of a control desuperheater or attemperator located i n either the location Of the Oil or water drum. The hing front burners is frequently dependent On the machinery (mart 'Onarrangement and may be in the Figures and indicate ventional), roof, or sidewall. some of these variations. In most i~tallationssome form of air heater is used with an economizer. The type and pmportions Of these auxiliw heat exchangers depend On the 'yd0 arrangement. If two stages of feed heating are selected, a steam air heater and an economizer are Often used'

'

81

BOILERS AND COMBUSTION

MARINE ENGINEERING

83

fig. 7 Two-drum, top-fired boiler with verfical superheater and ecanamizer

Ilg. 6 Tw-drum,

single-furnace bolkr with horizontal superheater

*I~lnlr would be incompatible with natural circulation. i'tco greatest disadvantage is the circulating pump Ib~lf,which is a potential source of trouble and mainCFII~IIOO. Fig. 5

Twdurnace, single-vptoke cantrolled superheat boiler

other furnace supplies heat to the superheater. Some designs incorporate a part of the superheater (called a primary superheater) in the reheater zone to provide additional protection for the reheater and to obtain the desired steam temperature characteristics [3]. The gas flowing from both the reheater and superheater combine in the main generating tube bank, and a single gas flow path is maintained through the auxiliary heat exchangers as in the single-furnace design. c. Forced-Circulation Boilers. Ever since the first boiler was used aboard ship, marine boiler designers have investigated and experimented with various means

to reduce the size and weight of boilers. A boiler arranged for natural circulation of the water and steam requires low waterside pressure drops which can only be obtained by installing sufficient downcomers and risers. This adversely affects size and weight. By supplying a pump to either augment or supplant natural circulation, a smaller and lighter boiler can be designed for a givengteam output (41. The circulation in such a boiler is said to be controlled or forced. The chief advantages of this are that very small-diameter tube^ with a high resistance to flow can be used in arrangements of heating surfaces and steam drum location^

'rlln LaMont boiler, shown schematically in Fig. 9, la rr typical example of the forced-circulation type. Wlllln wed abroad, it haa not found wide application in Ill@~rrrtrinefield in the U. S. The LaMont boiler uses a sltrgle clrurn into which the heating surface discharges a

wlato~ilatr of steam and water. The circulating pump e~teklnr~ ia supplied by gravity from this drum and forces refiller lllrough the generating tube surface, which is ~ I ~ ~ I ~ ! I Jof U Ia) number ~ of tube circuits arranged between r! hlslllbibutingheader and the steam drum. The inlet HI P~I!II tube is fitted with an orifice to balance the flow ~ralrrhnnoowithin the various circuits. This is necessary III 11l8bdnun adequate flow of water in each tube dependllrl ik oxpected heat input. The furnace, oil burners, sl~l~arl~nr~tor, and economizer are similar to those of $ 4 I ~111,ttl-airculation boilers.

Fig. 8

Slngle-furnoce, gas-bypass reheat boiler

BOILERS AND COMBUSTION

MARINE ENGINEERING

WATER INLET4

CONVECTION BANK RADIANT HEAT ABSORBINGSECT\ON

~ g 9.

Schematic of LaMant forced-circulafion boiler with economizer and superheater

of 150 to 300 psig. Boilers of this type are usually built only in small sizes and supply up to 7500 lb of saturated steam per hour. Because of the difficulties in maintaining feedwater chemistry, adequate water flow through parallel tube circuits which would be required for higher capacity boilers of this type, and the control of superheated steam temperatures, the once-through boiler is not well-suited for marine propulsion purposes. e. Supercharged Boilers. The superchased boiler has the characteristic of using combustion pressures higher than one atmosphere in the furnace to take advantage of higher gas densities and higher gas velocities than are available in the usual marine boiler. Figure 11 is a typical supercharged boiler. This unit is an outgrowth of the Velox boiler which has been used in a few stationary power plants for a number of years. In

I (11orlomisers of either the bare-tube or extended-surface t ~ v l ~ [we c r used to increase the temperature of the incoming

Iwlwuter by cooling the flue gases leaving the boiler. Ail- I~oatersare used to increase the temperature of the a~)~t~l)ustion air so as to promote better combustion of Iilln fuel. In the case of gas-to-air heat exchangers, air I ~ ~ n ~also c l r improve ~ the boiler efficiency by reducing 1,110 tomperature of the flue gases. By using low~ l r n ~ ~ u low-temperature re, exhaust or turbine bleed 3bnr1,rn to heat combustion air, as in the case of the steam dlu I~oeter, the overall cycle efficiency is improved. 'I'llcmo various types of heat exchangers may be used rrl~~yly or in combination with each other. a. Economizers. An economizer is a simple heat u~c\l~nt~ger consisting of a bank of tubes connecting an I ~ ~ l nr~nd b outlet header located in a relatively cool gas Iel111 mrature zone beyond the boiler main generating I~alllt. Supplied with water at a temperature near that Iuavil~gthe last feedwater heater, the economizer supplies ~rlrlihionalheat to the feedwater by cooling the flue gas. Irr lrlnrly installations the economizer is the final heat cttallttnger in the exhaust gas path. I t may, however, ko followed by an air heater where a higher efficiency is

typo# me forced circulated by the main feed pump. In $~l~nt*rkl, they are designed to heat the incoming feedwater CII willliinabout 35 deg of saturation temperature. They rre r~~rrangedfor counterflow of the water and the

The work of compression shows up, in good measure, as an increased temperature of the combustion air. As

A- FAN B-OIL BURNER WITH IGNITOR AND FLAME SCANNER C- FURNACE D-GENERATING COILS E- STEAM SEPARATOR F-STEAM TEMPERATURE LIMIT CONTROL

Fig. 10

Once-through boiler

d. Once-Through Boilers. The boiler in Fig. 10 is an example of once-through boilers used for auxiliary steam. Water is passed through the heating surface in one continuous circuit by the feed pump. The boiler is basically one long spiral tube arrangement composed of a economizer and a transition zone, where evaporation is completed, which surrounds the furnace. The feed pump pressure determines the outlet steam pressure, which may be 1200 to 1800 psig, dthough for the usual marine installation the pressure is in the range

tive naval vessels. The original Velox boiler, from which supercharged boilers evolved, was a forced-circulation boiler. However, subsequent supercharged units have employed circulation to avoid the extra complication of the circulation pump. f. .Waste-Heat Boilers. In vessels powered by diesel or gas turbine engines, the exhaust gases contain considerable available heat. Boilers placed in the stack to reclaim this otherwise wasted heat are called wasteheat boilers. Usually they generate low-pressure saturated steam which can be used for purposes such as tank heating, galley, and space heatingIf desired, they may be designed to bum oil when the main unit is shut down. Basically, waste-heat boilert3 consist of a bank of generating tubes that are either bare Or or of the extended-surface ~ Y P - Either forced circulation may be used. 1.3 Auxiliar), Heat Exchangen. In addition to the steam generator, several forms of a d i a ~ heat exchangers are inwrporated in boilers to impr0ve the efficiency and the overall operation of the plant

' h o nimplest economizer arrangement is the bare-tube !,up0 ~ ~ this n d was the form the first economizers took. 8flwuvcr, it was recognized that the use of extended ~ ~ l ~ ' f #tou oincrease the total heat-transfer surface for a PII lorlgth of tube would provide significant increases performance without penalizing weight and space rullwidorutions adversely. Figure 12(a)shows anefficient b41tn of extended surface in which flat studs are spaced rb dlimclogangles around the circumference and at %-in. IiikerfvaInalong the tube. 1h(tulldedsurface can also take the form of spiral fins #@ldpd on the hlbes or of cast iron Or alu~linumgill rings ~~~~1111~d 01 shrunk onto steel tubes as shown by Fig. 12(b). ba Air Heaters. The cooling of hot flue gases by the iilPo1rlillb! combustion air is one of the oldest of concepts Iily)r()ve boiler efficiency. In addition, heated air ~ I F U V ~ ~ ~an O B additional beneficial effect by promoting @@illd lblld complete combustion of the fuel. This can irn~ortancein the relatively small furnaces used III lr\~tl*ino boilers. Alr htraters fall into two broad classifications, the ke~~ll~~~!l'r~tive and the regenerative. In the recuperative bvlle, II(!IL~ from the products of combustion passes

1:

through a partition which separates the products from the air. Tubular and plate-type air heaters are examples of recuperative air heaters. In the tubular heater (Fig. 13) the walls of the tubes transfer the heat from the gas to the air. The plate-type heater is not c o m m o ~ yused in the U. S. in marine service. In it the air and gas are separated by plates through which the heat flows. In the regenerative air heater, heat is first stored in the structure of the heater itself .as it passes through the hot gas stream. The heat is then givenup to the air as the structure turns through the airstream. The air preheat& shown in Fig. 14 is an example of this type [7]. I t consists of closely spaced heating elements packed into a revolving frame. The frame speed is constant and is controlled by a small electric motor. The frame speed is selected such that the elements will absorb heat from the gas with a good temperature merential and, at the same time, 'the elements will heat the incoming combustion air to the highest possible extent. The upper section of the air heater is in the cold-air zone and also "sees" the coolest gas. It is usually arranged SO that the heat-transfer surface can be conveniently removed in easily handled s e ~ t i o ~ s - ~ a l l ~"baskets'-since d corrosion and fouling may occur there. These baskets

MARINE .ENGINEERING

is removed by steam traps. The latent heat of this steam which would otherwise be rejected in a condenser is returned to the boiler via the hot air. 1.4 Boiler Terms and Definitions. The. location of some of the more important boiler elements are shown in Fig. 5. For an understanding of marine boiler technology, a review of the applicable terms and definitions of various essential boiler parts may be helpful. The following terms and definitions are based on the standmds of the American Boiler Manufacturers Association

of a superheated vapor. Boiler hand.. . . . . . . . . . . . . . Boiler arrangement is described by reference to the location of the uptake gas outlet with respect to the designated front of the

Fig. 13 Tubular air heahn

Heating surface. . . . . . . . . . .

close all or e portion of a steam generator unit.

may, in addition, be provided with a ceramic coating similar' to porcelain enamel for protection against the

for treating the boiler water are introduced. Circulation ratio. . . . . . . . . . . The ratio of water entering a

are used as supply hbes to supply water to a drum or header. fercrad circulation. . . . . . . . . Circulation in a boiler by mechanical means external to the- boiler. Pllrnaoe screen. . . . . . . . . . . . One or more rows of tubes arranged across the furnace gas outlet. Pursl~cevolume. . . . . . . . . . . The volume contents of

may flow from the steam drum to the water drum or header. That surface which is exposed 'to the heating medium for absorption and transfer of heat to the heated medium, including any fins, gills, studs, etc. attached to the outside of the tube for the purpose of increasing 'the heating surface per unit length of tube, '

'

steam, usually expressed as the percentage by weight.

well system through which fluid flows downward. or box inside the steam

atural circulation. . . . . . . .

watertube boiler convection bank which is normally provided with a blowoff valve for periodic removal of sediment collecting in the bottom of the drum. Circulation of water in a boiler caused by the difference in density between. the water in the down. comers and the watersteam mixture in the gen-

BOILERS AND COMBUSTION

MARINE ENGINEERING

AIR IN

GAS OUT

1

I

t AIR OUT

89

I ROTOR

SEALS

I G A S IN

(a) Assembly

(a) Assembly of typical section

(c) Crimped spiral fln Fig. 15

(b) Replaceablebaskets Fig. 14

Rotary regenerative air heater with replaceable cold-end baskets

. . . . . . . . . . . The plates, centrifugal sepaRadiant heat absorbing.. . . . The projected area of tubes Steam baffling.. rators, or baffles arranged surand extended metallic surface (RHAS) to remove entrained watcr faces as viewed from the from the steam. furnace. Included are the --walls, floor, roof, and partition walls in the plane of the furnace exit screen. Steam or steam-and-. . . . . . . A pressure chamber located at, the upper extremity of II water drum Heat-transfer apparatus for Reheater. . . . . boiler circulatory system i t 1 heating steam after it has which the steam generateti given up some of its original in the boiler is separated heat in doing work. from the water and fro111 A tube through which steam Riser. . . . . . . which steam is discharged and water passes from an a t a position above a watcr upper waterwall header to level maintained therein. the steam drum. ~

Steam air heater

R1111~rlv)ater.. . . . . . . . . . . . . . A group of tubes which absorbs heat from the products of combustion to raise the temperature of the vapor passing through the tubes above the saturation temperature corresponding to its pressure. 'I'rr11yr311l;-tube wall. . . . . . . . . A waterwall. , in which the tubes are substantially tangent to each other with practically no space between the tubes. I'llllr I I I I I ~ ~.. .. . . . . . . . . . . . . A group of two or more rows of tubes forming part of a watertube boiler circulatory system and to which heat is transferred from the products of combustion mainly by convection.

Tube sheet. . . . . . . . . . . . . . . The part of the drum or header which the ends of the tubes penetrate. Unheated downcomer. . . . . . A tube not exposed to the products of combustion in which water may flow from the steam drum to the water drum or header. Watertube.. . . . . . . . . . . . . . . A tube in a boiler having the water and steam on the inside and the products of combustion on the outside. Water-cooled furnace. . . . . . . A furnace wall containing watertubes arranged to form a waterwall. Welded, mono-wall, or. . . . . A waterwall in which the membrane wall tubes are welded together (or to filler bars between them) to form a continuous furnace wall.

90

BOILERS AND COMBUSTION

MARINE ENGINEERING

Section 2 Consideratiofls in the Selection of a Boiler 2.1 General. Many factors influence the design and selection of steam generating equipment to produce the required quantities of steam at the design pressure and temperatures for a particular installation. Efficient operation when burning the various fuels available throughout the world is a requirement. The boiler also must fit easily and conveniently within a minimum of engine room space, yet be accessible for operation, inspection, and maintenance. Although light in weight, it has to be sufficiently rugged to operate dependably under adverse sea conditions. Operation over a wide load range, with a minimum of attention, and operating characteristics compatible with a high degree of automation are also required. The factors used in both the thermal and structural design must be conservative to provide assurance that continuous operation over extended periods of time will be provided with minimum maintenance. Finally, the boiler must meet the rules and regulations of the regulatory bodies. 2.2 Cycle Requirements. The design of a marine boiler is directly affected by the heat cycle selected by the ship's designer. Over the years steam pressures and temperatures for marine power plants have advanced by a series of broad jumps. After each jump there has been a pause to consolidate the gaina, review the operating results, and plan the next jump. In general, marine steam conditions have not advanced as rapidly as those in use ashore. In part this has been due to the relatively small horsepowers involved and in part ta the demands of the ocean environment. As the safety of the vessel and its personnel is dependent upon a reliable power plant, each new advance is made only after adequate experience is accumulated with the last. High steam pressures and temperatures may make reductions in the size and weight of a given propulsion plant possible, or permit a higher horsepower installation in the same space. During World War 11,most combat naval vesgels operated at 600 psig-850 F while steam to 450 psig-750 F was widely used in merchant ships. In the postwar era the Navy advanced to 1200 psig-950 F (nominal) for its combat vessel construction. In the late 1940,s and 1950's a significant number of merchant vessels appeared using steam at 600 psig-850 F and 850 psig-850 F. By the 1960's almost all new construction used 850 psig-950 F steam; several large vessels used steam (in some cases with reheat) a t 1500 paig-950 F. Machinery plants utilizing steam st pressures of 850 to 1500 psig and temperatures from 950 F to 1000 F are characteristic of most commercial steamships built during the 1970's. The quantity of steam produced by a marine boiler can range from as little as 1500 lb/hr in small auxiliary boilers to over 400,000 1b/hr in large main propulsion boilera. Steam outputs of 750,000 lb/hr or more per boiler are practical for high-power installations.

2.3 Heat Balances. The fuel cost per shaft horsepower is one of the deciding factors in establishing the characteristics of the boiler installation and whether or not the installation is economically sound. The fuel rate can be decreased by the use of higher steam pressures and temperatures or a more sophisticated cycle can be employed by the use of reheating, economizers, and/or air heaters, more stages of feed heating, etc. The designer must analyze these factors in light of initial cost, maintenance, weight, and space requirements versus the savings resulting from increased thermal efficiency. As steam pressures increase, it is essential to use additional heat-reclaiming equipment in the boiler unit. This is because of the corresponding increase in saturated ' steam temperature which results in a higher gas temperature leaving the boiler bank and thereby reduces the boiler efficiencyat a given firing rate. Reheating the steam improves thermal efficiency but requires larger boilers and special provisions to protect the reheater during astern operation. High steam pressures and temperatures, along with reheating, are more likely to be used in installations of 30,000 shp and up, where the value of the fuel saved may well justify greater initial cost and cycle complication. In addition, the utilization factor or load factor in such vessels is apt to be much higher, giving added impetus to the establishment of more efficient designs [9]. It is from the detailed heat balances prepared by the marine engineer that the quantities of steam and feedwater flow are determined. In the usual plant from two to four stages of feedwater heating are used to supply water to the boiler at temperatures from 270 to 400 F. Boiler efficiencies of over 90 percent are possible. However, to minimize corrosion and maintenance in the cold-end heat exchangers and uptakes, it may prove advantageous to limit the boiler efficiency to 88.5-90 percent with some fuels. Fuel oils vary widely in quality and often contain significant amounts of sulfur which can form sulfuric acid if there is condensation in the exhaust gas path. Corrosion and maintenance costs should be balanced against the possible savings in fuel costs derived from a higher boiler efficiency. 2.4 Fuels and Methods of Firing. The characteristics of the fuels which will be available to the ship in its usual trade should be established early in the design process. This will permit the optimum selection of equipment for burning the fuel and cleaning the boiler. In addition, a suitable selection of uptake temperatures and materials can be made for the entire boiler plant so as to reduce corrosion and maintenance problems. Most marine boilers are oil-fired, with wood, gas, and coal-fired boilers less common. Wood firing is generally confined to riverboats operating on streams with an abundance of nearby timber and is not an important

i11nl oxcept perhaps in some remote parts of the world. ( I~r~-fired boilers are used primarily on power or drill Imrgtrs which are fixed in location and can be supplied ~ I I I I I I~hore. At sea, tankers designed to carry liquefied ~iul,~rrr~l gas may use the natural boil-off from their cargo &#a lllulks as a supplemental fuel. This cargo gas I~~~iI-off is collected and pumped to the boilers where it is I~~rrncrtl in conjunction with oil. The oil burners serve BWpilots to provide ignition stability and also to augment tire l l t r l ~ available t from the gas. The quantity of boil-off ~vrilt~ble from the liquefied natural gas is a function of r t r ~ hiont r sea and air temperatures, the ship's motion, and It10 trnrgo loading, among other things, and may vary I r c ~ nduy ~ to day. C !old-lired boilers have persisted chiefly in older vessels trljer~~l~irlg on lakes and rivers, and in ferries, colliers, tti&dI t~ndtowboats operating in coastal services. Their t n ~ ~ l l l ~have o r ~ decreased steadily year by year as labor r114tw rino and air pollution control is expanded. M o ~ lcoal-fired marine boilers used hand or stoker n r i ~ ~ aThe . use of stokers, particularly the spreader Bylre, gormitted firing rates per square foot of grate l ~ ~ l r r u~tpproximately o 40 to 50 percent in excess of those tor ha~idfiring. This resulted in boilers which were far nlura aompact and lighter than those designed for hand I/glrrp; but even they were much larger and heavier than u(i4rsd boilers designed for comparable steam outputs. !3rllv~rizedcoal firing, widely used ashore, has seldom Rri~usud a t sea since the,furnace volume necessary for d@iii~m,Lmvel, low heat release rate, and satisfactory kmbuatian requires a tall boiler. The high fly ash kdllrg of the flue gas aggravates tube erosion, slagging, dtaak emission problems. Qilwwore used as boiler fuels as early as the 1870's but f#d nos aohieve widespread use until the automobile age fgqulrecl a world-wide petroleum industry. Compared &$ ei,har fuels, oil is easily loaded aboard ship, stored, lnbroduced into the furnace; and the firing equiplVequireslittle costly maintenance. The small l$#i@unCof ash and contaminants it contains does not mdre t,ha extensive ash handling facilities required for

ma6

@&jl flrlng,

1) ~lrelrldbe recognized that fuel oils from different WIFOPH, while similar in heating value, have varying r n ~ u n b aof contaminants which may be harmful in

mpiew ways. The major contaminants consist of @@a of vfinadium and sodium. As a class, they are ~ l e A"a~h"and their presence must be fully taken into l@@@~irt by the designer. Likewise, the sulphur content wry over a range from almost none to as much as &f psroallt in "sour" crudes; sulphur has a decided en the cycle efficiency which can be obtained @t!t&rb tierious corrosion in the economizer, air heater,

a&&

~ptrtlees.

Tkr oompounds of vanadium and sodium affect the 11

af the superheater. If oils to be burned in a

trtde are especially rich in these constituents, r sriperlla~tttorcan be designed with tube metal temCinltricrr lower than normal to avoid the possibility

91

of severe slagging and tube metal corrosion problems. Cold-end heat exchangers designed with full recognition of the sulphur content present in the fuel will experience a minimum of corrosion and expensive maintenance. A boiler designed to take advantage of low-cost residual fuel oils can always burn lighter fuels if the situation justifies it. However, a boiler with tightly packed heating surfaces designed for light oils such as diesel or aviation turbine fuels would not perform satisfactorily on residual fuels for very long. Gas-side fouling and oil burner and combustion problems in the furnace could be anticipated. 2.5 Effect of Ship Delign and Other Machinery on Boiler Design. Factors such as space, weight, and the

requirements of the regulatory bodies are major considerations in the design of a boiler. In addition, however, the prospective vessel owner or his naval architect may have preferences regarding the boiler design and specific design requirements. These preferences may include the number of boilers, types of boilers and their arrangement, locations of major connections, the use of economizers and/or air heaters, fining, and evaporative ratinga, and the type and method of firing. Life-cycle costs can have a bearing on the preference likewise, since the total cost and labor involved in maintaining a previous design or construction may be reflected in the owner's specifications and result in the selection of an improved design and construction. a. Space. The space provided for the machinery is held to a minimum by the naval architect because the space occupied by the machinery produces no revenue. The boiler designer is usually required to adapt the boiler design to the available space. The boiler height may be limited by deck or machinery casing locations. The fore-and-aft or depth dimension of the boilers may be controlled by bulkhead locations, access, or tube renewal space requirements as well as the location of control consoles, main engines, etc. To a large extent the aviilable space determines the economy of the design. A height restriction is particularly serious, since it usually necessitates increased boiler width or length to obtain the required heating surface. This generally results in a marked increase in boiler cost, weight, and the base area occupied. b. Weight. With drum-type boilers, the minimum ~ efficiency is obtained with rninimuql weight f o maximum furnace depth, maximum tube length, and the maximum number of tube rows. Limiting the height may restrict capacity because of reduced circulation. It may also result in tube slopes and in burner clearances less than the minimum necessary for a good design. In header-type boilers the width is changed by increasing or decreasing the number of header sections, and the height is varied by changing the number of tube clusters in a header. Because of reductions in the number of boiler sections and the length of the steam drum, it is readily evident that long, narrow, and high boilers lead to minimum weight. Further, since the maximum efficiency for a given heating surface is obtained with the

MARINE ENGINEERING veatest numb& of tube rows in height, header-type these limits may be modified in the special specifications boilers always should be arranged with the maximum issued for a particular class of vessel. m he Maritime height, rnmimum length, and minimum width which are Administration follows a somewhat similar procedure and usually establishes evaporative and furnace heat compatible with the design conditions. The minimum weight of any type of boiler will vary release rates for each design. considerably with desi@ conditions;increases in evapora- . 2.6 Boiler Design Criteria. heo ore tical and practical tive rating, burner capacity, or air pressure decrease the considerations have led to the establishment of boiler weight of a boiler design@ for a specified steam output. design criteria in a number of areas not directly associated With a fixed evaporative rate per square foot of heat- with the regulatory bodies' rules, which concern mainly absorbing surfaoe, the weight of a boiler per pound of pressure-part scantlings and construction techniques. generated will be less for boilers with greater steam The design criteria are most important in the areas of output, since certah boiler parts remain fixed in size and combustion, heat absorption rates, circulation, and pressure drops through the boiler system. They provide weight over a reasonable range in capacity. Weight is greatly dependent on space also. Generally the yardstick by which various boiler designs Can be the larger the physical dimensions of a boiler for a given compared for their suitability for specific applications. a. Combustion. At the heart of a successful boiler output, the greater its weight. is a properly designed furnace and fuel burning systemThe ocean environment is no place Regulations. to test unproven principles. This became evident in the If the fuel supplied to the furnace is not burned cleanly construction when it was and completely within the furnace throughout the range early days of recognized that some rules and regulations were necessary of operation, it will not be possible to accurately predict to protect life and property. These rules were not the performance of the evaporator-superheater comintended to inhibit the designer or innovator but rather bination. For example, the total steam generated may for comparison of be insufficient, the steam temperature may be incorrect, to provide a sound basis and or the efficiencymay be lowered by incomplete combusnew designs with older successful designs. tion Or improper excess air. Disastrous boiler explosions, common to both marine A number of criteria by which combustion in furnaces and stationary boilers, resulted in the establishment of a boiler inspection senice and strict regulations can be gaged and by which different furnaces Can be care, and operation of compared have been developed. In"general1 with the governing the steam boilers. In the design of marine boilers the exception of the furnace heat absorption rate which is applicable regulatory rules and standards must be rigor- derived from the actual heat transfer calculations ously followed. Most units built for American-flag developed for the furnace, they are empirical relationthe requirements of the United States C o ~ t ships with little theoretical value; however, they can be ships used to compare similar boiler designs provided their Guard and the American Bureau of Shipping. Boilers for naval combatant ships are built in strict limitations are recognized. The criteria most fI'equent1~ used for these comparisons accordance with Navy specifications, although for are: auiiliary naval vessels the use of the United States Coast Guard or the American Society of Mechanical rate per cubic foot of furnace volume. Heat Engineers codes often is permissible. For foreign-flag ~ i rrate i per ~ ~square footof radiant heat absorbing ships, the rules and regulations of other midatory surface. bodies would apply. In addition, many shipyards and Heat absorption rate per square foot of radiant heat operators of large fleets have established their own absorbing surface. supplementary rules and regulations. Since the requirements of the various regulating and A brief review of these factors will sewe to indicate their inspection groups differ, specifications must be clearly importance and usefulness. The heat release rate per cubic foot of furnace volume defined to assure fabrication and installation of boilers which will be approved by the boiler inspectors. Fur- is useful in comparing geometrically similar furnaces, ther, it is important that all competitive designs be to the but while widely used because of its simplicity, it is not A design difference caused by the an important criterion. The heat released is the product same me of inapplicable specifications could be the deciding of the hourly fuel rate and its higher heating value, factor in final cost or wei&t evaluations, particularly ignoring any heat above 100 F in the combustion air. If radiant heat absorption rates, furnace gss temperaon high-pressure unito where a difference in pressurepart thickness might involve not only price and weight, tures, and furnace tube metal temperatures are satisfactory, the only limitation on the heat release rate Per but also design and fabrication changes. ~~~t rules pertain to const~ctionand the inspection cubic foot of furnace volume should be that imposed of materials, and establish very few by the ability of the firing equipment to maintain good and The use of a high1 yet satisperformance limitations. ~ l t h o u g hNavy specifications combustion conditions. furnace volume heat r f ? l ~ ~late e peat1y factory, rates per cubic foot of furnace limit the heat installation of high-capa~ib~ lightweight facilitates the volume, per square foot of radiant heat absorbing surface, and per square foot of total heating surface, boilers in a minimum of space.

BOILERS AND 'I'll() temperature within a boiler furnace can be ~llilll~rolled to a large extent by the effective radiant heat r~lno~~bing surface (RHAS) present in the furnace [lo]. 1 IPIIII is radiated from the flame envelope to thee heat t~lluorhingsurfaces with the uncooled refractory surfaces n i \ l ~ i as r ~ ~an intermediary, receiving heat from the flame ru~dl111cnre-radiating most of the received heat back to 1 II* ll~uneand cold surfaces. For a given heat input or ~ ~ " rate, I I K the heat absorbed per unit area decreases wlIllr ILILincrease in total RHAS. The greater the RHAS ~ I I * ~t'aaterwill be the total amount of heat absorbed by (Itn Fllrnace. Therefore, the temperature of the gases \

COMBUSTION

93

boiler has more demands placed on i i than a comesponding shoreside boiler. In addition, the heat input and the steam output of the marine boiler are probably higher than for a comparable application ashore. It is customary to consider a momentary roll of 30 deg from the horizontal and a momentary pitch of f5 deg when computing static and dynamic loads. In establishing circulation, boilers are u s p d y designed for a permanent list of 15 deg and a permanent trim by the bow or stern of 5 deg. The latter, when coupled with the momentary pitch of 5 deg, means that in the fore-and-aft direction, the boiler may be as much as 10 deg from the horizontal. The arrangement of the tubes and steam-

,

.. .

n

f

MARINE ENGINEERING boiler must likewise be capable of prolonged periods of steady operation a t its design rating. Also, in port it may be subjected to long periods of operation at low or minimum outputs. Cleaning, with the exception of the daily use of the mot blowers or occasional attention to the atomizers in the oil burners, is normally deferred to the annual or biannual period when the vessel is in a shipyard for other maintenance. This must be fully taken into account by properly locating soot blowers so they are effective; by using the optimum burner combinations for the range of fuel-oil types anticipated to be bunkered; and by using the best possible arrangements of economizer, air heater, boiler furnace, and generating surfaces to pinimize fouling. must also include margins in the scantlings The ---- desim of tubes, supports, casings, and other parts exposed to corrosive flue gases or waterside contaminants. Simple and easily accomplished maintenance procedures can also do much to assure that the boiler will be available to meet the ship's requirements. The duty cycle may also have a pronounced effect on the number of boilers selected. A single boiler may be employed in ships of up to about 90,000 shp. Two or more boilers may be selected for higher power levels or where redundancy is desired or required. Single-boiler vessels have proven reliable in service and should continue to do so. This is in part due to the fact that a boiler kept continuously in service reaches thermal equilibrium and can have the waterside chemistry optimized. In general, from a boiler performance point of view, the least number of boilers which can deliver

the required steam will prove to be the best selection for any particular vessel. e. Automation. Widespread use of automatic controls @ndmonitoring equipment has made bridge control of the power plant possible and has permitted a reduction in the number of watch-standers in the machinery space. These desirable improvements have added additional -considerations .to the problem of designing a suitable boiler. Of prime importance is a fuel burning system that can respond rapidly throughout the range of operation from standby to maximum power without a fireman's attention. It must do so to prevent excursions in steam pressure and reduce water level fluctuations (shrink and swell due to changes of the volume of steam present in the boiler), which might result in water carry-over into the superheater [12]. Burners can be designed to operate over the full boiler range with all burners in service, or other burner types with less range can be sequenced, that is, placed in or out of service on command by the control system. Suitable flame-monitoring safeguards and purge interlocks are necessary in varying degrees of complexity depending on the extent of manual supervision desired. Feedwater regulators, steam temperature controls, d a t a logging equipment for flows, pressures, temperatures, levels, etc. are all available from the simple to the ultrasophisticated. The owner and his naval architect usually select the scope of equipment and advise the boiler designer so that the boiler and burner combination can be made compatible with it. See Chapter 21 for additional discussion regarding automation and controls. -

atttl in part on the space available for the installation ant l its operating requirements. 'I'ho quantity -of fuel required is determined from the ~ltwirod steam generator efficiency, the given steam prtrnHure, temperature, and flow, the feedwater temperaI,II~'o, and the heating value of the fuel. 'I'ho fuel characteristics and quantities establish the' I~lrlburning equipment to be employed. This in turn ICI~H the excess air requirements. Combustion calculal l l r l r l ~are next made to determine the hourly quantities rlf llue gas flowing through the unit. The exit or stack baa tomperature to which the flue gas must be cooled b nohieve the desired efficiency is determined (Fig. 16); R I I ~if experience indicates that it is attainable or otherw l ~ t r natisfactory, the design can proceed. If not, a~rr~t~hor selection of efficiency must be made and the ealaulations repeated. 'I'ba furnace exit gas temperature is next calculated. Ell@ value is dependent on the radiant and convection 11ewt-transfer surface installed in the waterwalls, floor, tr~nf,t~ndscreen (radiant only) as well as the extent of refractory present. Next, the gas temperature drops &acl tho heat absorbed by the screen and superheater are dsbarmined. The size and spacing of tubes and the &mount of surface are assumed initially. These are lhrn modified to provide the desired steam temperature rrild cronservative tube metal temperatures as necessary. V~uallyseveral screen and superheater combinations are Invemtigated to determine the most economical solution. r heater surfaces ke gas temperast outlined, initial aterials for tubes,

the heating surfaces established, the draft loss all components is calculated. If the draft loss the capability of the fan desired, the heat drafts previously calculated are adjusted he tube spacing, number of rows crossed or height of the boiler components.' A ers may be necessary ce of draft require-

Section 3 3.1

General. The fundamental boiler design prob-

lem is to determine the proper proportions of the various heatrabsorbing surfaces to use the maximum heat available in the products of combustion. A proper design will accomplish this at the lowest cost on a lifecycle basis. Each component must be integrated with the other elements of the unit to provide a balanced design in which the first costs and fuel, maintenance, and operational costs will be a minimum over the useful life of the ship. In no way must safety or reliability be compromised by these cost considerations. For the steam generator system, the following must be considered :

1. Fuel burning equipment 2. Furnace 3. Boiler generating surface 4. Superheater (and reheater if used) 5. Economizer and air heater

6. Attemperator (or control) and auxiliary desuperheaters 7. Circulatory and steam separator system 8. Casing and setting 9. Cleaning equipment 10. Safety valves and other mountings 11. Feedwater and treatment 12. Foundations and supports 13. Combustion air supply system 14. Uptake gas duct system and stack These considerations require many interrelated steps. In most cases, a number of assumptions must be made in order to initiate the design. ks the design calculations proceed, the assumptions are refined to achieve the desired accuracy in the final analysis. The first step is the selection of the basic type of boiler, superheater, and economizer or air heater (or both) to be used. This selection is based in part on preference

95

BOILERS AND COMBUSTION

I

drops of water and steam through all comm the economizer feedwater inlet to the superuted. They, in turn, estabeconomizer design pressures tho safety valve settings. A circulation analysis @aprepared using the heat absorptions determined . From this, the bes are adjusted as for each design. er can make very ntially reduce the

@,P

Fuel Combustion. The basic function of a

ilrp frirnace is to generate the maximum amount of rrb Imm a given quantity of a specific fuel. A useful

RAOlATlON AN REFERENCE 0

FOR UNITS WITHOUT STEAM AIR HEATER

STbCK GAS TEMPERATURE, F

Rg. 16

Efficiency v* stack gar temperature

secondary function is to generate steam in the furnace wall tube circuits. The theoretical aspects of combustion have been well known for many years. However, the achievement of good combustion within the furnace of a relatively small marine boiler requires practical knowledge and experience. Complete combustion can be obtained provided there is sufficient time (a function of furnace volume), turbulence (provided by the geometry of the burner assembly), and a temperature high enough to provide ignition. Combustion may be defined as the chemical combination of oxygen with the combustible elements in the fuel. The common fuels have only three elemental constituents which unite with oxygen to produce heat. The elements and their compounds, as well as their molecular weights and combustion constants, including heating values, are given in Table 1. Oxygen combines with the combustible elements and their comgounds in accordance with the laws of chemistry. Typical reactions for the combustible conatituents of fuel oil, based on the assumption that the reaction is completed with the exact amount of oxygen required, are : for Carbon (to COa) for Hydrogen (to HzO) for Sulfur (to SOa)

+ + +

++ +

C 0 2 = COZ AQ 2H2 0 2 = 2Hz0 AQ 2s 302 = 2508 A Q where A Q is the heat evolved by the reaction. The heat evolved or heat of combustion is commonly called the "fuel heating value" and is the sum of the heats of reaction of the various constituents for one pound of the fuel considered. The heating value of a fuel may

96

BOILERS AND COMBUSTION

MARINE ENGINEERING

1113 calculated from theoretical considerations or may be clt!l,ormined, for an actual oil, by burning a sample in a I)olnb calorimeter (see Chapter 23 for additional discusi4o11in this regard). 111 testing fuels by a bomb calorimeter to determine the l l t r ~ b tgiven up, two values may be reported: the higher ([)I' Kr088 Or upper) heating value and the lower Or net Il~~~ltling value. For the higher heating value, it is nafl''med that any water vapor by burning the I1,Vtlrogen constituent is d l condensed and cooled to the l11ll~i1~1 temperature in the calorimeter at the end of the tsrl,. The heat of vaporization, about 970 Btu/lb oil, is inoluded in the reported heating value. For the lower ~isrtl1iug value, it is assumed that none of the water vapor mnclo~~sesand that all the products of combustion vermin in a gaseous state. In the United States higher I ~ e ~ t ~ vdues i n g are used as they are available directly fl'c~lllthe calorimeter determinations and because of the @stnll>li~hed practice of buying fuel on a higher heating vnlue basis. The lower heating values are generally ~irreclia European practice. Fuel Analysis. For design and comparative IrlitlptrNos, the standard reference fuel oil is #6 fuel oil [@uelrur C) having the following characteristics [13]:

CHEMICAL COMPOSITION (percent by weight) Carbon 87.75 Hydrogen 10.50

Total

100.00

tho following expression :

By weight By volume

%OXYGEN%NITROGEN 23.15 76.85 21.00 79.00

The rare gases are included as part of the nitrogen constituent. Air is assumed to be supplied to the forced-draft fan at a temperature of 100 F, a rklative humidity of 40 percent, and a barometric pressure of 29.92 in. Hg. Under conditions air has the following physical prope*ies: Dry-air density, lb/cu ft MoistureJ lb/lb of dry air Mixture density, lb/cu ft Specific heat

0.0709 0.0165 0.0701 See Fig. 3 of Chapter 2

Based on the foregoing fuel and air standardsJ analysis will show that the s~ic-,iometrical or theoretical quantity of dry air to burn one pound of fuel is 13-75Ib. From this, the following quantities of air for various excess percentages are determined : Excess air, percent Dry air, It, Moisture, Ib Moist air, lb Volume, cu f t (at 100 F, 29.92 in. Hg) dry air moist air (40% RH)

0 5 10 15 20 13.75 14.44 15.13 15.81 16.50 0.23 0.24 0.25 0.26 0.27 13.98 14.68 15.38 16.07 16.77 194 200

204 210

213 220

223 230

233 240

The ultimate analysis of the fuels actually encountered in service varies from that of the standard reference fuel. Figure 17 shows the effect of these variations on

98

MARINE ENGINEERING

BOILERS AND COMBUSrlON

MARINE ENGINEERING Table 2

Oil Burner Clearances

PARTIAL STUD TUBES

FULL STUD TUBES TYPICAL STUD-TUBE WALLS

wider angle is employed to shorten the flame length and produce a wide bushy flame while a narrower angle increases flame length and decreases width. The burner manufacturer should always be given the opportunity to review the projected furnace design so the best possible installation can be obtained. Generally suito' able burner clearances are shown in Table 2. When firing Bunker C oil, it is customary to use the minimum clearances established by experience. These may be Fig. 19 Change in efficiency vs. load decreased perhaps by six inches, if distillate oils are fired. Furnace depths of watertube boilers which are front-fired are usually limited to a minimum of six feet boilers the large amount of fuel and air to be introduced although there are highly rated boilers in service with into the furnace necessitates a multiple burner instal- furnace depths of only five feet. The selection of the oil burner must also include the Each size burner has a minimum rate of operation type of atomizer to be used. There is a wide variety of below which it becomes unstable and there is risk of atomizers from which a selection can be made. The losing ignition. In part this is a characteristic of the alternatives include: steam atomization (internal mix), burner, but the forced-draft, fuel, and control systems steam mechanical (external mix), straight mechanical, also have an influence. The minimum rate is of great return flow, rotary cup, and others. Of these types, the a much simpler plant results when all internal mix steam atomizer has the greatest turndown importance burners can be left in service at all times. When in and provides the smallest and most uniform particle port or during rnanuevering conditions, the minimum size over its wide range of operation. Development0 oil flow capability must be less than that required by have materially reduced the quantity of steam required the plant demand, if frequent safety valve popping or (80 to 120 lb/hr-burner depending on the maximum oil steam dumping is to be avoided. Both of these actions capacity) so that earlier objections to the loss of evaporated water have been more than offset by the other waste steam and lead to increased maintenance, Burner sequencing can be used effectively to follow advantages. The uniform and finer article size has the load demand where burners with limited range or provided more surface area for combustion of the fuel lower higher-than-desired minimum flows are used. Solid- droplets. This has permitted less excess air and necesstate, computer-controlled logic systems are often used draft losses since the high air velocity to sequence burners; hovbever, this equipment canincrease sary to provide the turbulence to burn larger droplets i~ no longer required. costs considerably [15]. The number of burners selected usually results in a Care must be taken in arranging the burners to provide for even air distribution to each burner within burner draft loss equivalent to about 35 to 50 percent of combustion with a minimum of the total draft loss of the boiler unit. The burner draft the windbox to varies with the volumetric flow of air through it. excess air. The clearances between the burners and the loss At any given air flow, a change in the temperature of tho to prevent interference furnace walls must be air will increase or decrease the draft loss in the ratio of The furnace volume must be large and impingement. enough to provide the time necessary for complete the change of absolute temperatures. In desiping a to take place before the gases enter the super- boiler with an air heater, it is standard ~racticeto limit heater screen. Satisfactory combustion has been ob- the air temperature leaving the air heater and enter in^ tained at furnace release rates of up to 1,500,000 the burners to no more than 600 F and refer ably l e s ~ to assure long life and prevent overheating of the burher Btu/cu ft in marine boilers. If the preliminary design ~ i e l d san excessive air parts. Each burner manufacturer has his own recommended temperature, the designer must reapportion the surface*, clearances and the shape of the flame can be adjusted to possibly adding a small economizer, to reduce the air This is some extent to modify them when necessary, done by changing the spray angle of the atomizer. A heater air outlet temperature to an acceptable value.

TANGENT TUBES

MEMBRANE WELDED TUBES TYPICAL BARE-TUBE WALLS

Fig. 20

TUBE AND TILE

A

Furnace wall construction

1 02

BOILERS AND COMBUSTION

MARINE EN

estimates of furnace exit gas temperatures were not necessary because of conservative firing rates and the use of saturated steam. Those units which generated superheated steam usually had several rows of boiler tubes between the superheater and the furnace. Consequently, a large error in the calculated furnace exit gas temperature had very little effect upon superheater performance. In units with superheaters located dose to the furnace, however, the furnace exit gas temperature must be determined accurately to assure a satisfactory superheater design. In addition, an accurate determination of the heat absorption in the various furnace waterwall areas is necessary to provide adequate water circulation with a practical number of supply and riser tubes. When estimating the furnace gas temperature, most designers use formulas based upon the Stefan-Boltzmam law, which states that the heat absorbed by radiation is proportional to the difference between the fourth powers of the absolute temperatures of the radiating bodies and receiving surfaces (see Chapter 2). However, in a boiler furnace the exact determination of radiant heat transfer, or heat absorption, is extremely complex and depends upon: the furnace size and shape; the radiant beam (mean distance from the radiating gas mass to the absorbing and the re-radiating surfaces); the partial pressure of the products of combustion; the amount, type, and effectiveness of the heat absorbing surfaces; the ratio of the heat absorbing to the refractory surfaces; the type, quantity, and heat content of the fuel; the amount of excess air; the temperature of the combustion air; the latent heat losses; the emissivity of the various surfaces and the radiating mass of gas; and the flame luminosity. Designers usually calculate furnace exit gas temperatures and heat absorptions by rational methods and then, as a check, plot the calculated values against empirical data derived from boiler tests 121. b. Radiant Heat Absorbing Surface. In evaluating the radiant heat absorbing surface, the flat projected areas of the walls and tube banks are used. The spacing of the tubes in the boiler bank adjacent to the furnace has no effect upon the furnace temperature; but with widely pitched boiler tubes, a large percentage of the radiant heat is absorbed in the tube rows behind the furnace row. Furnace waterwalls and roofs usually consist of bare or covered tubes (Fig. 20) and, with the exception of bare tangent tubes or welded walls, the effectiveness of the absorbing surfaces is less than the black-body coefficient of 1.0 considered for the furnace rows of boiler tubes. The furnace gas temperatures usually are not accurately estimated in preliminary analyses since the general design characteristics are of primary interest, and an approximate estimate of furnace gas temperatures and heat absorption rates can be made with knowledge of the boiler and the firing conditions. Thus, with the assumed excess air, the heat content of the products of combustion and the adiabatic temperature can be determined. Further, the approximate furnace size

provides an indication of the water-cooled surface8 and estimates can be made of the surface absorption effectiveness and the expected furnace gas temperature. In approximations of this nature it is usually desirable to estimate both the furnace temperature and the heat absorbing surface on the low side when firing oil. This increases the estimated furnace heat absorption and assures a margin of reserve in the final design. However, with coal firing it is more important to estimate the furnace gas temperature on the high side to preclude the possibility of operating with furnace temperatures above the initial ash deformation temperature. In a boiler furnace, both the furnace exit gas temperature and the heat absorption can be changed appreciably, for a given firing rate, by varying the amount of radiant heat absorbing surface. The furnace gas temperature and heat absorption also can be lowered, at any firing rate, by increasing the excess air (Fig. 21), except when operating with a deficiency of air. The additional air increases the weight of the products of combustion per pound of fuel fired. This decreases the adiabatic temperature since there is less heat available per pound of products of combustion; and, as indicated by the Stefan-Boltzmann law, lowering the radiating temperature reduces the heat absorption rate. Generally, the radiating temperature is assumed equal to one third of the adiabatic temperature plus two thirds of the furnace exit gas temperature. c. Heat Absorption Rates. The furnace heat absorption rate per square foot of radiant heat absorbing surface increases with larger heat release rates. However, the percentage of the total heat released which is absorbed in the boiler by radiation decreases with an increase in firing rate, and varies from as much as 50 percent, or more, at the lower firing rates to about 15 percent at the higher firing rates; see Fig. 22. This results from the fact that the adiabatic temperature remains practically constant, except for changes due to variations in excess air and combustion air temperatures, over the entire range of boiler operation, while the temperature of the gases leaving the furnace and entering the tube bank increases with the firing rate. Even though the furnace heat absorption rates may be conservative, the furnace exit gas temperatures may be excessive with respect to ash fusion temperatures and slagging. This is true particularly in coal-fired boilers where the gas temperatures entering the tube bank should be less than the initial ash deformation temperature. Because of the lower ash fusion temperatures of oil slags, they pass out of the furnace in a gaseous or molten state and are not amenable to control by reducing the furnace exit gas temperature. They must be considered in the design of the superheater. d. Tube Metal Temperatures. In boilers, the heattransfer rate across the boiling water fdm on the inside of the tubes may be as high as 20,000 Btu/ft2-hr-F; however, when estimating tube metal temperatures, a transfer rate of only 2000 Btu/ft2-hr-F is usually assumed in order to provide a margin against the resis-

103

EXIT-WITH 15%

FIRING RATE,PER CENT OF FULL OUTPUT FIRING RATE,PER CENT OF FULL OUTPUT

.

I I l k c t of excess air on odlobotic tind furnace gar temperature

Flp. 22 Relotianhip of rodlon) heat absorption ond Aring rote

ratings, including port loadings. However, at t.he same time they should not be so high as to cause high casing temperatures or excessive furnace maintenance. Because of the requirements for exceedingly lightweight and compact units for naval installations, evapntly, with a steam pressure of 600 psig orative ratings in naval boilers are 3 to 4 times greater steam temperature) and a heat input than those common to most merchant installations. Consequently, the furnace exit gas temperatures in the full-power to overload range are about 2800 to 3050 F when firing oil with approximately 15 percent excess air. Adiabatic, or theoretical, flame temperatures are about 3450 to 3500 F with oil firing, 15 percent excess air, and 100 F combustion air. With combustion air temperaappreciable and it is good design practice t o tures of 300 to 350F, the adiabatic temperatures L tolerance for variations in the quality of the increase to approximately 3650 to 3700 F. Although furnace heat release rates vary considerably, practically all oil-fired merchant boilers are designed for heat release rates of 65,000 to 125,000 Btu per cubic foot of furnace volume per hour at normal rating-approximately 15 to 20 percent of the corresponding full-power heat release rates on naval boilers. The heat release rate per square foot of radiant heat a b s o r b i surface is generally in the range of 200,000 to 250,090 Btu per horn on merchant boiler designs.

104

-.

MARINE ENGINEERING the minimum longitudinal tube pitch (direction parallel to the drum and perpendicular to the gas flow) consistent with good manufacturing practice and acceptable drum design, unless the draft requirement or the type of fuel fired dictates the use of a greater pitch. Manufacturing and fabricating practices permit the use of +-in. metal ligaments between 1-in. or la-in.-OD tubes. The circumferential, or back, pitch (direction parallel to the gas flow) of the tube usually is set to maintain circumferential or diagonal ligament efficiencies2 equal to, or better than, the longitudinal ligament efficiency in the drums. Tube arrangements utilizing a minimum back pitch reduce the drum periphery required for a given number of tube rows and allow the use of smallerdiameter steam drums provided the steam drum release rates are satisfactory. With such arrangements, the size and weight of the boiler can be reduced. When designing for high steam pressures, it is often necessary to increase the tube spacing in order to improve the ligament efficiency and reduce the thickness of the drum tube sheet [l6]. If this is not done, large thermal stresses may be set up in the tube sheet. It also i~ possible to maintain close tube spacing and yet reduce the drum tube sheet thickness by using tubes with the ends swaged to a smaller diameter. The number of tube rows installed should be limited so that an impractically large steam drum diameter i~ not required and so that heat absorption in the last tube rows is adequate to maintain good circulation. The tube length should be such that the total absorption per tube does not result in too high a proportion of steam it1 the water-steam mixture leaving the upper end of tho tubes. b. Header-Type Boilers. Single-pass header-typo boilers (Fig. 3) generally have two rows of 2-in. t u b e ~ above the furnace and if-in. or 1-in. tubes in tho remainder of the bank. In these boilers a group or cluster of fourteen 1-in. tubes can be substituted for ono of nine la-in. tubes. Thus, in boilers having the samo width, length, and number of tube clusters in height, 25 percent more heat absorbing surface can be installed by substituting 1-in. for la-in. tubes. However, tho advantages resulting from the compactness of the 1-in.tube boiler must be balanced against the greater tolerance provided by the la-in.-tube boiler for poorer feedwatcr quality. For the new header-type boilers that arc1 installed, chiefly in motor vessels for auxiliary steam purposes and in drill barges and dredges, the feedwator quality is apt to be such that the selection of larger tubo sizes will offer more reliability. c. Boilers Delivering Superheated Steam. Practically all marine boilers built recently deliver superheated steam from convection-type superheaters. In these boilers, the generating tube bank is arranged in two

Naval boilers are designed for ratings four to five times greater than those used for merchant marine boilers. Radiant heat absorption rates vary greatly depending upon the firing rate and the amount of cold (watercooled) surface in the furnace. Generally, a radiant heat absorption of 120,000 Btu per square foot of cold surface per hour is considered satisfactory for continuous overload operation of merchant boilers with treated evaporated feedwater. This results in an absorption of about 100,000 Btu per square foot of cold surface per hour at the full-load rating. There are merchant boilers in continuous service with radiant heat absorptions of approximately 150,000 Btu per square foot of cold surface per hour; and most naval boilers have been designed for radiant heat absorption rates of 150,000 to 200,000 Btu per square foot of cold surface per hour at overload rating, but operation a t this rating is infrequent. 3.4 Boiler Tube Bank. The arrangement of the boiler tbbe banks is established after development of the preliminary furnace size. The simplest type of tube bank is that of a boiler delivering saturated steam. Usually two sizes of tubes are used in such banks. The tubes in the rows adjacent to the furnace absorb considerably more heat than those in the other rows and, therefore, should be of larger diameter to increase the water flow. The total heat input to the furnace row tubes is the sum of the radiant and convection heat transfers; in general, the convection heat transfer is approximately 5 to 20 percent of the radiant heat transfer. This relatively wide range in convection heat transfer results from variations in tube diameter, tube pitch, gas mass flow rate, and the temperature difference between the products of combustion and the tube surface. The number of tube rows installed is primarily dependent upon the circulatory system and the desired gas temperature leaving the tube bank. The gas temperature leaving the boiler tube bank varies with changes in steam pressure, firing rate, and tube size and arrangement (the tube arrangement may be either staggered or in-line). However, sufficient boiler heating surface must be installed to obtain exit gas temperatures which result in economical operating efficiencies and do not require excessive stack and breeching insulation. Generally, the exit gas temperatures should not exceed 750 F unless economizers or air heaters are used. The resistance to gas flow can be varied appreciably in drum-type boilers by changing the pitch of the tubes in a direction perpendicular to the gas flow. This change is not possible on header-type boilers because of the fixed tube pitch and, therefore, variations in resistance to gas flow must result from changes in boiler width, tube length, and the number of tube rows. a. Drum-Type Boilers. Mbderately rated drumtype boilers usually have 13-in. tubes in the furnace roes, but these are increased to 2 in. in boilers of higher rating. One-inch and I&-in. tubes are common in the 2 Ligament efficiency is the relative strength of the ligamenln There is no standard pitch for tubes between main tube adjacent tube holes in a drum or header as compared with in drum-type boilers. However, it is customary to use a drum or header having no holes.

I

BOILERS AND suc~(~ior~~. The section between the furnace and the +!~~l~n~~ltoater is known as the "waterscreen" and the other F~UI~~~IO installed II, beyond the superheater, is called the " l ~ ~ ~ i l rbank" ir. or "generating bank. " 'I'l~ti~ i z eand arrangement of the waterscreen greatly r1l;fecrln the design of the superheater. A superheater I~~c~~iiCtsl d.oser to the furnace behind a few rows of widely j~ll.irl~n(l tubes in the waterscreen provides a relatively ili~l, nl,nl~mtemperature characteristic over a wide range ~ r l1-abi11g since the radiant and convection heat-transfer titten tmd to .complement each other. However, a i ? ~ i l l ~ r I ~ ~located t ~ t e r farther away from the furnace i.uiiat,inn behind a deeper waterscreen has a steam i r ~ ~ ~ l ~ n ~ characteristic ature which rises steeply with ~IIIIIQ@B~MO~ rating, due to the greater effect of convection 11111LIIN reduction in radiation heat-transfer rates. Navril boilers usually have waterscreens consisting of t l l i ; ~or :four rows of tubes and merchant marine boilers ,

FIRING RATE. PER CENT OF FULL OUTPUT

Fig. 23

de a relatively constant steam de range of rating. Superheaters. The superheater must deliver the ed ateam temperature during the operating life t during the initial trials or test cted performance must be mainvariations in firing d excess air. The necessity of unscheduled~oqtaiges rder to maintain performance.

Temperature characteristics of radiant and convection superheaten

these two factors and the surface. Increasing the temperature differential takes advantage of the available temperature potential, while an increase of the heattransfer coefficient necessitates a larger resistance to gas flow. Full advantage should be taken of a high temperature difference, but the entering gas temperature should not be so high as to result in excessive tube metal temperatures or high-temperature fuel ash corrosion (these are primarily a matter of location). The change in steam temperature with firing rate should be a minimum in order to prevent excessive temperatures during maneuvering and, again, this depends upon location. Steam velocities should provide for good distribution of steam, minimum tube metal temperatures, and acceptable steam pressure drops; all of which require correlating the effects of size, location, and the arrangeure dictates the thickness of the super- ment of the steam passes. which in turn is an important factor in the a. Types and Characteristics. The radiant and of superheater pressure drop and tube convection-type superheaters are the two basic types. They are, as their names imply, superheaters which receive heat by radiant or convection heat transfer and they may be arranged horizontaily or vertically. In the radiant type the steam temperature decreases with increased rating since the quantity of heat absorbed by radiation does not increase proportionally with steam flow; see Fig. -23. In the convection type, the steam temperature generally increases with increased rating are designed to have a because the heat absorption, due to greater heattransfer coefficients and higher inlet gas temperatures, urface can be obtained by increases a t a faster rate than the steam output. hcnt-transfer coefficient and the CemperaMost superheaters are a combination of the two basic oducts of combustion and types in which the designer builds in a radiant combsorbed is the product of ponent to achieve a flatter temperature characteristic.

II 1

MARINE ENGINEERING

106

(a) Three-pass hairpin loop type

BOILERS AND COMBUSTION

1 07

(b) Two-pass continuous loop type

~ i24~ Schematic . arrangement of hairpin and continuous-loop superheaters

GAS TEMPERATURE- F

,

108

MARINE ENGINEERING

Economizer elements (particularly the extended-surface type) are more expensive than boiler tubes. In air heaters, part of the advantage resulting from the improved temperature difference is offset by the high resistance to heat flow across the air flm [17]. Therefore, the proportions of component surfaces must be studied carefully to obtain the most economical overall arrangement. The minimum temperature of the feedwater to most merchant marine economizers vaxies between 270 and 280 F. The standard feedwater temperature for most naval installatior~is 246 F. This lower temperature is satisfactory because a premium fuel with a low sulfur content is used. Since the gas temperature leaving the economizer cannot be less than the inlet water temperature, .it follows that high feedwater temperatures limit the obtainable efficiency. Consequently, with high feedwater temperatures, economizers are not often used unless they are installed in conjunction with air heaters. I n an air heater, the minimum uptake gas temperature is dependent on the entering air temperature. Therefore, the attractiveness of air heater installations is due to the possibility of operating with a high boiler efficiency when using feedwater temperatures in the range of 300 to 450 F. When steam turbines are bled for regenerative feed heating,'the plant efficiency is increased about 1 percent for each 100 deg F rise in feed temperature due to the reduced heat loss in the condensers. Whether this improvement in efficiency warrants the expenditure required for additional feed heating and other equipment should be carefully weighed for each application. The use of an air heater necessitates an increased air pressure to the boiler unit because of the additional resistance to air flow through the air heater. Air pressures also must be increased when using economizers because of the relatively high resistance to the gas flow across the economizer, but, for boilers of the same size operating at comparable firing rates, an air heater installation will usually require a higher total air pressure than will a unit fitted with an economizer. w:1R rlg. IU. Air heaters are not pressure vessels, so the tubes can be If the uptake gas could be cooled to a temperature fabricated from mechanical tubing (less expensive than equal to the steam's saturation temperature by the use of pressure tubing) that is lightly expanded into the tube an infinite amount of heat absorbing surface, the improved efficiency would only be 83.75 percent. There- sheets. However, economizers are part of the pressure fore, air heaters or economizers must be installed to system and must be designed to withstand the main increase full-load efficiencies to the 88-90 percent range feed pump discharge pressure, to operate without leakage, usually desired. Further, the use of high evaporative and to withstand thermal shock. a. Air Heaters. Increased efficiency and reduced ratings a t any given steam pressure increases the need boiler maintenance can be obtained by improving comfor additional heat reclaiming equipment. When air heaters or economizers are installed, the bustion. Preheated air can improve combustion, reduco proportions of the boiler, air heater, and economizer boiler sooting, and reduce the possibility of ignition loss surfaces must be balanced. Usually, the temperature particularly at the extreme low end of the firing range. Practically all of the older marine air heaters were of differential between the products of combustion and the the tubular type; the regenerative types were not often heat absorbing fluids in the economizer and air heater is used. However, in recent years, particularly for highgreater than that in the last section of the boiler tube bank. This is advantageous in reducing the heat powered installations, the rotary regenerative air preheater has found wide application. A typical absorbing surface required for a given heat recovery.

Experience has shown that the diligent use of sootblowing equipment (particularly mass-action retractable units) usually can keep superheater surfaces satisfactorily clean for a year, or more, of opelation and that manual cleaning and washing of the external heat absorbing surfaces are required only during scheduled overhauls. h. Reheaters. The design of reheaters involves the same procedures and considerations that are pertinent to superheater design. However, the steam distribution and tube metal temperature problems are more critical since reheaters must be designed for exceptionally low steam pressure losses if a high cycle efficiency is to be obtained. Steam or combustion gas can be used as the heating medium in reheaters. When steam heating is used, the temperature of the reheated steam usually is limited to 550 to 600 F, since it is customary to use condensing rather than superheated steam as the heating medium because of the much higher rate of heat transfer. The use of gas reheaters is necessary if high reheat steam temperatures and cycle efficiencies are required. Such reheaters may be fired separately or installed in the boiler proper. Separately fired reheaters are not common because they require an individual firing aide and renewal clearances. as well as additional piping, -. controls, breechings, firing equipment, fans, etc. 3.6 Air Heaters and Economizers. Air heaters and/or economizers are necessary to obtain high boiler efficiencies. Preference alone should not arbitrarily influence the selection of either since the design of the power plant and it? performance characteristics greatly affect the choice. The temperature of saturated steam at a pressure of 850 psig is 528 F and the temperature of the products of combustion leaving the boiler tube bank would be, for a conservative boiler design, approximately 150 deg F above this value, or about 675 F. When firing oil, and operating with 14.0 percent COs in the products of combustion (approximately 15 percent excess air), this uptake gas temperature would result in an operating efficiency of only about 80 percent as can be see" from

BOILERS AND COMBUSTION

nxt~rnpleof a regenerative air heater is shown in Fig. 14. I l,n gastight casing forms part of the boiler forced-draft rir nnd uptake gas ducts. The heater is separately ~llourltedabove the boiler and suitable expansion joints mBoused in the ducts joining the two [7]. 'I'l~eessential component of the heater is the rotor in wllitill the heat-transfer plate elements are packed. The aila for combustion is passed axially through one side of Llln lutor while the flue gas is passed through the other aliltr in the opposite direction. As the rotor turns, heat I. nitltinuously transferred from the gas to the heating r~l~*lrbco; heat is also continuously given up to the air as the lioated plates traverse the air side. Counterflow I I tl~e ~ gas and air insures efficient heat transfer. 'I'l~n heat-transfer elements are made of corrugated and flnl alloets which are alternately packed in the main secI11111 of the heater and in the cold-end baskets. The coldel111basket is designed to be readily removable for cleanilly or replacement when conditions warrant. For daily elon~irig,a cleaning device consisting of a mass-action ar~ol~ blower is installed. Air and gas bypass dampers fiiw ctn integral part of the preheater and are useful in rature sion at imize soot ers can be made air heaters are of the horizontal type vertical type is no%often used since it is stall considerably more surface for a eat absorption than would be needed for the e, it is customary he gas across the In the vertical type the gas usually passes the tubes and the air crosses the tubes. ontal tubular air heaters generally utilize in-line &rrangements. These facilitate cleaning of the far more r heat transfer

isite tubes aximum the heatacross and y as about and, thus, ses with a s, both the tube size satisfactory, in most bes with *-in. tube de of the length of ow, the number of

109

tube rows, and the number of gas and air passes. This facilitates determination of the heat-transfer rates and the heating surface. The preliminary assumptions are then adjusted, if necessary, so that the surface arrangement and heat transfer provide the required heat absorption. Gas and air flow patterns also must be analyzed since maldistribution could reduce heat absorption, increase fan power, reduce or elevate tube metal temperatures, or restrict the capacity of the boiler unit. Air heater designs are usually predicated upon inlet air temperatures of 100 F, and exit air temperatures ranging from 300 to 450 F at the normal full-load operating rate. Design exit gas temperatures of 290 to 320 F are common for tubular air heaters and result in boiler efficiencies of 88.5 to 88 percent. Regenerative air heaters can be designed for lower uptake gas temperatures for a given risk of corrosion since for the same air and gas temperatures the heating surface metal temperature is somewhat higher than that of the tubular heater. Gas temperatures from 240 to 260 F are common for regenerative air heaters with boiler efficiencies of 90 to 89.5 percent respectively. Both the weight of the gas produced and the specific heat of the flue gas are greater than that of combustion air. Therefore, when firing oil with about 15 percent excess air, the reduction in the temperature of the products of combustion passing through the air heater is about 13 percent less than the rise in air temperature. In air heaters the heat-transfer coefficients across the gas and air films are of about the same magnitude, and high resistance to heat flow is encountered in the gas film on both sides of the tube. b. Economizers. Marine economizers can be grouped into two general classifications, the "bare-tube" and the Uextended-surface"types. They are generally nonsteaming and are usually arranged for counterflow of the water and the products of combustion. This results in larger temperature differentials, and greater heat absorption can be obtained. The counterflow nt permits a higher boiler efficiencybecause the temperature can approach that of the inlet omizers use tubes ranging in size from to 2 in. arranged in the form of either hairpins or continuous loops. The hairpin type consists of U-bend tubes that are welded, or expanded, into headers. Single or multiple rows of loops can be used as well as two or more headers. I n the continuous-loop type, each tube element consists of a length of tubing bent back and forth to form the desired number of rows; the ends of the tube are attached to the inlet and outlet headers, usually by welding. Since only two headers a& required, the number of tube joints is greatly reduced as may be noted from Fig. 26. There are many types of extended-surface economizers. The most prominent are those having steel studs or circumferential fins of aluminum, steel, or cast iron (see Fig. 12). Features common to all extended-surface

110

MARINE ENGINEERING

cient varies as the 0.65 to 0.70 power of the gas mass flow rate. Usually, if the economizer width is increased, a reduction can be made in the height of the economizer. Most marine economizers use counterflow arrangements with up-flow gas and down-flow water. The water pressure drop at about 25 percent of the normal full-load operating rate should be equal to, or greater than, the static water head in order to prevent recirculation. This minimum pressure drop requirement is not necessary if parallel-flow, up-flow gas and water, nonsteaming economizers are used, since the water pressure in the outlet header always will be less than that in the inlet Multiple water passes are often used in hairpin-type economizers to obtain satisfactory water velocities and pressure drops. These arrangements have both counterand parallel-flow relations between the water and the products of combustion, and the calculated heat transfer should be based on the average of the flow arrangements. Most continuous-loop and extended-surface type economizers have a single water pass arranged for flow counter

I

111

BOILERS AND COMBUSTION

bare-tube economizers the temperature drop across I ~ I I tube I ~ wall is small and, for all practical purposes, the Iltll)(j rmtal temperature can be considered the same ( J I I L ~ of 411'

the water it carries. Tube metal temperatures

extended-surface elements also are about the same as

Idltl ldjacent water, although the tip temperatures of the @xl,trrrdedsurface are considerably higher. I't'udence, and often regulations, requires a check valve 11) Illlo connecting piping between the economizer and the fltflfirndrum to prevent the loss of steam pressure in the ~Vnlltof an economizer casualty. Further, the valve fanilitates filling the economizer, particularly since a wnh@rhead of several feet is required to lift the check if k l i troonomizer ~ is located above the normal water level, h bypass line around the economizer will allow rrl8arrition of the boiler with an economizer outage. Ihbwover, few economizers are fitted with bypass lines k~aarlaeof their cost and the piping r ~ o m p l i c a t ini~~

Fig- 27

Drum-type desuperheater

BOILERS AND COMBUSTION

MARINE ENGINEERING

rlrr~rr sections and there is a definite transition zone I~ebwoonthe heated downcomers and the riser tubes, the 111t~alion of which varies considerably with changes in I IIH boiler firing rate. 111 the U-tube analogy, there is initially a vertical ja"mure plane a t the bottom on which the pressures ~ ~ n r l , oby d the hot and cold water legs are equal. As

NOZZLE

u(111111pressure plane in the lower water drums, or Ir@a(lurs,the pressure corresponding to the flow of water Ilrrough the downcomers is equal to the product of the I~oatlof water and its density minus the resistance to flow. 'I'lrk pressure must balance the product of the head of wnt1trrin the risers and its density plus the resistance to

+

N~TE IN ATYPICAL BOILER (SEE FIG 44) THE SIDEWALL AND ROOF CIRCUIT A,AND THE SCREEN AN0 FLOOR CIRCUIT B ARE SUPPLIED BY TWO DOWNCOMERS C. FURNACE FRONT AND REAR WALLS D AND GENERATING BANK E ARE SUPPLIED BY HEATED BANK DOWNCOMERS F. A HEAD TO WATER-STEAM MIXTURE F L W CURVE IS REQUIRED FOR EACH INDEPENDENT CIRCULATING SYSTEM. AND WOULD BE SIMILAR TO

TF

= '6

VENTURI-MIXING AND THERMALSLEEVE SECTION

Fig. 29

Fig. 28

Uncontrolled and controlled steam temperature

External-spray desuperheater

ture to the design value. The temperature of superheated steam is a function of rating and for the usual marine boiler rises as shown by the "uncontrolled curve" in Fig. 29. To make the most effective use of the materials in the superheater and main steam piping, the final steam temperature can be controlled so as not to exceed the design value. This can be accomplished by passing a portion of the superheated steam through a desuperheater in the drum. The location of the outlet and inlet connections is usually "interpass"; a typical arrangement is illustrated in Fig. 30. The desuperheated steam is returned to the last passes of the superheater where it mixes with the main flow to deliver the design temperature [2, 3, 61. A manually operated valve or an automatically controlled valve is used to regulate the temperature at all rates above the "control point9' (that point on the uncontrolled steam temperature characteristic curve which crosses the desired controlled temperature line). 3.8 Circulation and Steam Baffler. The natural circulation characteristics of the boiler and the type of steam drum b a a n g are determined after the arrangemerit of the heat absorbing surfaces has been established. Generally, because of the effect of the steam drum baffles upon the circulatory system, simultaneous analyses are made. circulation calculation procedures are in part empiricaland in part theoretical. The purpose of the is to establish a system of downcomers, riaers,

(a) Interpass, across restricted pass ~ g 30 .

[b) External bypass, three-way valve system

Interpass control desuperheater arrangements

and generating tubes which will insure that each tub0 receives an adequate supply of water in relation to the maximum heat absorbed. a. Circulation: Boiler Tube Banks and Furnace Waterwalls. The circulation characteristics of f u r n ~ o waterwalls and boiler tube banks are determined by tho same procedure and, since the water-steam ratio decreases with increased rating, the characteristics must be established for the maximum contemplated rating. In analyzing boiler cirqulation, it can be assumed that each circulating system is, in effect, a U-tube [6, 181. The riser section of the U-tube is that portion of the tubn bank in which the flow of steam and water is upward a* heat is applied. The downcomer section consists un" heated tubes or those ~ o r t i o n sof the tube banks ill which the heat absorption is considerably lower than "l the riser section. Because of the difference in fluid densities, heated tubes can act as downcomers for thf'

~ B @ Cdensities,

minus the riser friction 1oss-a quantity as the net available circulation head [3]. 111 most circulation analyses the steam geaerated in the rlmr tubes is calculated and the water-steam flow, as well r ~ t lthe net available head, is then determined for V L P ~ O ~ water-steam ~R ratios. In analyzing circulatory @hrrl.noteristics,it is customary to graphically plot both llro downcomer friction losses and the net available sirolllntion heads for the assumed water-steam mixture %ewu, As shown by Fig. 31, the flow tit which the ~ltr~dlt~ble head minus the resistance to flow through the ~ C e ~baffles nl equals the resistance to downcomer flow is that required to balance the circulatory system. From F ~ flows P at the balance point the percentages of steam by wlnnlo at the top of the riser tubes can be calculated. Tho percentage of steam by volume a t the top of the @@el'llubes must be such as to preclude overheating of ih@tlibes. If the quantity is excessive, the circulatory 6YPb111 must be redesigned to provide additional downkmflrrr, or the size and contour of the downcomers must )I atlonged to reduce the resistance to flow. It also CW&y ho necessary to change the location, size, and &llllt~llr of the boiler tubes to redistribute the heat ~Brrerptionand reduce flow resistance. f 11 a satisfactory circulatory system, an adequate @moullCof water must be supplied for each pound of ~ C ~ a gonerated. nl Therefore, if the percentage of steam b,Y vol~lmeat the exit of the riser tubes is used as a design @rlk@rlori, it is necessary to vary the allowable percentage MI @Irapressure changes since the percentage of steam by vtllulno will increase as the pressure is reduced because @f tlla irlcreased specific volume of the steam. Naval Btrllera nro usually designed for water-steam ratios (i.e., k@l#irll of water/weight of steam passing through the #elrornLiag tubes) ranging between 5.0 and 10.0, and hlel'bll~ttltunits usually fall in the range of 15.0 to 20.0 IC blra overload rates of operation. Lower water-steam p ~ b l onro ~ used on naval boilers in order to reduce the Crjlltlr ~ i a o and weight by minimizing downcomer bt~trwtl

I

Fie. 31

Characteristic head venus water-steam mixture flow for circulation calculations

b. Heated Downcomers. If evaporative ratings are conservative and the gas temperatures leaving the boiler do not exceed about 750 F, the first several rows of tubes will function as risen with the remainder serving as heated downcorners. As the firing rate increases, the high-temperature zone moves farther back into the tube bank and additional tubes become risers while a corresponding lesser number act as downcorners. If the firing rate is further increased, the number of downcomers becomes inadequate, circulation is impeded, and tube casualties may occur; when design analyses indicate such circumstances, external or unheated internal downcomers must be installed. c. External and Unheated Internal Downcorners. With conservative evaporative ratings, external downcomers 'are required for only those portions of the boiler in which the tubes cannot act as downcomers (i.e., a single tube row forming a furnace boundary, a shallow tube bank installed between two furnaces, or tube banks shielding a superheater from two furnaces). If downcomers are required for the main tube bank, they usually are located external to the tube bank even though the arrangement requires longer boiler drums. The use of unheated internal downcomers minimizes the drum length and eliminates tubes in the main boiler bank; however, unheated internal downcomers usually enter the steam drum at high water levels and they may lose water during heavy rolls or inadvertent reductions

STEAM OUTLET

upon the natural separation of steam and water. For higher boiler ratings a positive means of steam separation is required and compartmenbtype baffles, Fig. 32(b), are frequently used. Centrifugal steam separators are used primarily in highly rated merchant and Navy drum-type boilers; they are particularly desirable for boilers subjected to rapid maneuvering, fluctuating water levels, or high solids concentrations in the boiler water. Centrifugal steam separators may be arranged either horizontally or vertically in the steam drum as in Fig. 3 2 ( ~ ) . The resistance to flow through centrifugal separators is

(a) TRIPLE PERFORATED PLATE BAFFLE

115

BOILERS AND COMBUSTION

MARINE ENGINEERING

llloat merchant units having 48 to 54-in. drums and most llnval boilers using 46 to 60-in. drums. As power levels Ill(:rease,60 to 72-in.-dia drums are used more frequently 101) provide the necessary room for steam baffles and to lw()videthe capability of accommodating the shrink and n w d l that occurs when maneuvering. b- Headers. Headers for water walls or economizers ~ 1 . ousually fabricated from pipe stock. &llow forgings 1lltU' also be used especially for superheaters. They may b~round or forged to a rectangular or other cross @fl(ltion to facilitate tube installation. Tubes are lll~lulledby expanding or by welding. Htandard boiler and economizer tubes are fabricated from either electric resistance welded or seamless stock. t1:lo0tric resistance welded tubes are less expensive and lllbvo been proven to be as dependable as seamless tubes 111 boilers and economizers. Superheater tubes are made

ULATING FIREBRICK INSULATING BLOCK ASBESTOS CEMENT

!TEAM OUTLET

supplied to the downcomers is greater than that of the "frothy" water-steam mixture discharged from perforated-plate and compartment-type baffles. e. Effect of Drum B d e s on the Circulatory System. The steam-water flow through the steam drum baffles is in series with all of the flow circuits in the circulatory system. Thus, if the flow through one of the circuits is increased, for example, by the installation of additional downcorners, the flow through the steam baffles also is increased. This imposes an additional resistance in the circulatory system with the result that the flow in downcomers will not increase in direct propodion the additions made.

emperature to 130 F or less. Local areas, for where superheater inlet or outlet nozl;les

( C) CENTRIFUGAL SEPARATORS

FIBERGLASS

ng. 32 Typlcal steam reparation equipment

st boilers are of double-casing comtruction. An

boilers, combustion gases are discharged into ry space in the event of a leak. ( 0 ) BANK AND UPTAKES

materials vary. to suit the application; those for a particular unit can be readily determined. ral or strength members of the casing are used art some of the loads of the pressure parts. The ing bank and screen and furnace walls are eolf-supporting; however, the casing may lend these pads during rolling and pitching of the it is not on an even keel- It is U S U ~ 00 h o s u ~ ~ o r t tsuperheater he headers and the super' t'lbfis (wholly or in pad) , as well as the economizer l'lrlbtcr Or tubular), on the casing structLlre.

'

Fig. 33

Typical sections of boiler casings

Suitable access and inspection doors are required and their location is an important practical aspect of casing design. Provisions must also be made for differential ~expansion between the pressure parts and the casing and between the casing and the boiler foundation and surrounding decks, platforms, piping, etc. In large boilers where welded walls are used, another

>

"

,

>

rF

%

'h

T L

.*.t. \ *

MARINE ENGINEERING

BOILERS AND COMBUSTION

117

MARINE ENGINEERING

BOILERS AND COMBUSTION

MARINE ENGINEERING

give the operator a direct view of a light source which shines through the boiler uptake and the combustion gases. Another type employs a photoelectric cell and provides a readout on a meter scale calibrated in smoke density units; it may also be fitted to sound an alarm when a certain smoke density is reached. d. Instrumentation and Controls. The need for operating instruments and manual and/or automatic controls varies with the size and type of equipment, the method of firing, the proficiency of the operating personnel, and the desired degree of automation. Chapter 21 covers the application of control equipment to ship's propulsion plants. For safe operation and efficient performance, information is required relative to the water level in the boiler drum; burner performance; pressures of the steam and the feedwater; temperature of the superheated (md reheated) steam; pressures of the gas and air entering and the leaving principal components; feedwater and boiler water chemical conditions and particle carry-over; operation of feed pumps, fans, fuel burning, equipment; relationship of the and fuel actual combustion air passing through the furnace to that theoretically required for the fuel fired; temperatures of the water, gas, fuel, and air entering and leaving the principal component parts of the unit; and feedwater, steam, fuel, and air flows. Icor many years, marine boilers have been equipped with control equipment permitting steady operation at sea with little operator participation except while maneuvering. However, the trend is toward complete automation of the boilers so that, with the exception of starting up, they can be operated throughout the full range from standby to full load without manual adjustTo attain fully automatic operation, the development of adequate control components and system designs is essential. The operating characteristics of the principal and auxiliary items of steam-generating equipment must be fully known since these characteristics affect the degree of controllability, the scope of the controls required, and the response obtained. These in turn affect the safety of the installation and establish its economic justification. As an example, where the burners have a range of oper~tionor turndown capability equal to or greater than that required by the boiler, the necessity to sequence burners (or take them out of service) is eliminated. This, in turn, eliminates many decisions and functions that would otherwise be required of an automatic burner management system, and a simpler system may be selected. The degrees of control which can be achieved, in ascending order of sophistication, are manual, local supervised manual, remote supervised manual, automatic (nonrecycling), and automatic (recycling). These various types of control can best be delineated by relating their functions to burner operation. With the manual type of control, Fig. 40, a burner is manually purged and ignited. I t may be automatically

modulated but it is stopped manually. Although no operator function is ~erformed automatically, widerange burners can be used with automatic comb us ti or^ controls to facilitate dock-to-dock operation without manual participation. However, without boiler and burner monitoring devices, the operator must remain in close proximity to the boiler to provide the necessary surveillance. In the local supervised manual system, Fig. 41, a burner is purged and manually ignited, but certain ~ r o c e d u r e ~ and conditions are supervised by safety interlocks. ~ l l manual functions are performed and checked by tho operator a t the burner station during normal operation, and if the demand for steam is within the capability of the burners, unattended boiler operation is attained. Monitoring and safety interlocks are ~rovidedto alter the operation if an unsafe condition develops, and to trip the burner and/or the boiler, if necessary. After 11 trip-out, the operator must take the necessary correctivn action to clear the interlocks and recycle the burner and/or the boiler. The remote supervised manual system, Fig. 42, allow^ a burner to be purged and ignited by a ~ushbuttonor selector switch, modulated automatically, and securcd by a remote manually actuated pushbutton or selector switch. I t also provides supervision of procedures by safety interlocks. The burner is mechanized and all operating functioris are ~erformedby mechanical device^ initiated from a remote control station which indicaton whether or not each function has been performed correctly. This system of control does not relieve thtr operator of burner manipulation. He must devote hin undivided attention to the step-by-step procedures folstarting and securing burners, which is a time-consuminlr, process. This control system can only be justified i l l installations where the turndown capabilities of tho burners do not match the turndown requirements of tho boiler, and, the burners must be manipulated to covclr. the operating range. Its application will not meet thtr USCG requirements for an automatic boiler. The automatic (nonrecycling) control system, Fig. 43, involves a burner which, when actuated manually by 11 pushbutton, is purged, ignited, and modulated automatically; and although secured either automatically or remote-manually, the burner does not recycle automatically. When start and stop sequences are manually initiated from a remote control station, each function i l l the start-up and stop sequence is performed and checltatl automatically and all ~roceduresand conditions arc' supervised by safety interlocks. Since the operator may be required to initiate the start-up and securing of n burner to meet load requirements, this control systenl does not meet the USCG definition of an automatic! boiler. With an automatic (recycling) type of control systenr, a burner is purged, ignited, modulated, and stopp(-tl automatically, and the burner recycles within a prescribnd load range. 3.12 Sample Design Problem. The steps followcxl

BOILERS AND COMBUSTION

Fig. 40

121

Burner operation-local

manual control

Hlltrthm pressure, drum, approx.. . . . . . Hll(rfbmpressure, superheater outlet. . .

nuperheated 185,520 Ib/hr (losuperheated 16,870 lb/hr Pntdwater temperature. . . . . . . . . . . . . 41:Hiciency (based on 13% radiation auld unaccounted for losses and 15% flxcess air) . . . . . . . . .'. . . . . . . . . . . . . P'ud total heating value (standard h n k e r C 4- added heat in air). . . . 19,264 ~ t u / l b Pll(!l required. . . . . . . . . . . . . . . . . . . . . 14,349 lb/hr Alr temperature, leaving steam air houter . . . . . . . . . . . . . . . . . . . . . . . . . 811'flow (16.07 Ib/lb oil at 15% excess

nir) . . . . . . . . . . . . . . . . . . , , . . . . . . . . Ylue gas flow = 244,937, say. . . . . . . .

245,000 lb/hr

fpol' the example, only one rate of operation will be lalsul~btedalthough for an actual boiler design it is not ilfitl@llfil to calculate three or more rates to establish @l*kl'fi()~Ori~ti~ Curves of performance. Rated power will orl(lulated since this establishes the design meeting khr~ np(>eifiedefficiency and steam temperature. The h m l - ~ ~ s fdata e r are derived from the cumes and pro@@tlrlr.ao of Chapter 2. Boiler Layout. Two oil burners will be used to @MPP~,Ythe total oil flow of 14,349 lb/hr at rated power Mia ahout 8000 lb/hr each at overload. The necessary @!@#r~lces for burners of this capacity are obtained from the ~(rlocted burner manufacturer. Based on this !ltfl1rllll~tionand experience, an approximate furnace and bll@l' l h ~ o u is t prepared (see Fig. 44) from which the hlfllfitf(fvolume and heating surfaces can be estimated. k~. Furnace Calculations. The furnace volume, cold

*

@ul.f#fltr,and

radiant heat abs~rbingsurface @HAS) are

Fig. 41

Burner operatiolr-local rupewhed manual conko]

MARINE ENGINEERING

BOILERS AND COMBUSTION

To determine the shape emissivity factor, FBFA,the following data are required: VF = 2655 fta ST = 1200 fta S, = 1175 ft' PF = 1 atmosphere

X. = tube equivalent thickness =

k

Thotefore the firing density is [see equation (31) of

Fig. 7 of Chapter 2 the concentration factor -K la 0,086. The mean radiating length is L = 0.6q2655 = ft. Equation (31) of Chapter 2 can now be evaluated &odotermine the flame emissivity

123

=21 Dolog, O-D Di

2 -22 log. = 0.182 1.67

= tube conductivity = 310 Btu/hr-ft-F

Next, by estimating the corrected furnace exit ternperature, T E ~to, be 2200 F the tube film temperature can be approlrimated as

R I I ~from

EF =

~

~

~of furnace ~ exit~ temperature i ~ and furnace ~ tabrorpriar i

~

for an &/ST value of 0.98, FEFAis determined to be bSd4 from Fig. 6 of Chapter 2. In order for the calculation to proceed, it necessary w u m e several values of the furnace exit temperature. bibking this assumption

* ture and heat absorption can be calculated (See Sectio" 2 of Chapter 2)) based on the following furnace surfacO

The temperature coefficientf~ [equation (40) of Chapter 21 then becomes

f~ = 0.00003875T,r

0.95(1 - e- (o.o6a)(i)(s.a) ) = 0.353

FURNACE TEMPERATURE TE,

Ag. 45

= 1653 F

+ 0.1035 = 0.1675

With a flue gas flow W nof 245,000 lb/hr and two burners having 2-ft thmat diameters, the flue gas weight flow rate

G is 39,000 1b/ft2-hr. Since the furnace depth D is 14 ft, the surface heat-transfer coefficient hRw can be computed from equation (39) of Chapter as ~0.6a

,/&)a; equation (36) of Ohapter 2; Btu/ft2-hr

2,200 2,660

2,300 2,760

2,400 2,860

95,100

88,000

81,300

haw =

f r = 13.2 ~tu/hr-ft2-F

e246,000~o.ai4 TB~ = 2239 F

The adiabatic sensible heat in the combustion can be computed from equation (37) of chapter a fuel lower heating value of 17,500 Btu/lb and a fuOl Ensible heat of 46 Btu/lb (100 deg F rise at 0.46 heat), for pedect combustion the sensible heat bemmo* determined by the methods of Section 2 of Chapter T & R Bulletin 3-14 [lo] to be: Furnace volume = 2655 fta Projected surface = 1200 fta RHAS = 1175fta

Or

With a fuel higher rating value of 18,500 B t u m the furnace ratings at rated power are: Release rate =

LHV

QTAI

-

+ q~ 4- (ta - ~o)CPR

17,500

+ 46 + (278 - 80) (0.2445) (13.98) 13.98 4- 1

TUBE TRANSGAB VEEBE BACK FLOW

DIAM-

mture in the furnace is From Fig. 2 of chapter 2 the adiabatic flame tempertLture, TAt,is found to be 3990 F or 4450 RWith 15 percent excess air

3.5

2.5

100

850

\

641 F at drum saturation pressure of 975 psia

The screen, superheater, and generating bank performance calculations may be conducted as follows:

126

MARINE ENGINEERING

This practice is expected to becO1'lr' used after proper treatment (19, 201. In essence, this from corrosion. common, particularly at higher steam Pressures an(( entails: the removal from the raw water of those con1 stituents which are known to be harmful; supplementary single-boiler installations. Filming amines introdl~(~~sl treatment (within the boiler or connected system) of into the feedwater or steam lines also provide ~rotec1,l~)ll impurities to convert them into harmless forms; against corrosion, but by forming a coating on the mrlftll and systematic removal, by blowdown of boiler water surfaces rather than by changing the PH of the watts. of 4.3 Boiler Water. Boiler water is treated within I ~ I I I . concentrates, to prevent excessive boiler to prevent corrosion, the fouling of heat-absorblll# solids within the unit. surfaces, and the mntamination of steam. T h i S r e q ~ i ~ " ~ ~ The ultimate purpose of feedwater and boiler water the injection of chemicals into the steam drum W I I ( ~ I ~ treatment is to keep the internal surfaces free from deposits of scale or sludge and to prevent the corrosion they react with the residual impurities in the feedwi~ln-I of these surfaces. Hard-scale formations, formed by Properly controlled, internal treatment can mai111.nlll certain constituents in zones of high heat input, retard boiler water conditions within satisfactory limits [6, 1x1 an al1c:~li111~ Corrosion is minimized by maintaini~l~ the flow of heat rnd raise the metal to higher-than111 boiler water and this condition is usually expressed temperatures. This can cause overheating and The PH of w;lt1'l 'STP,lb/fta . . . . . . . . . . . . . . Volumetric speclfic heat at STp, I#1u/fta-F. . . . . . . . . . . . . . . . . . . . . . . lltrl~~tive heat transfer coefficient ('ompared to He for same gas (,ur?peratureand same power output 1tt31atlvepumping power compared to 110 for same gas temperature and mne power output. . . . . . . . . . . . . . . . lbflhtivepumping power compared I,o He.. .......................... lfdlbtlve cost of gas per IOOO fta at STP Italative total activity. . . . . . . . . . . . . . . llflll~tivegamma activity. . . . . . . . . . . . .

g..

are sealed to prsvent a loss of fuel or The fuel fission products to the reactorcoolant under all normal *he fuel elements are rnsembled in fuel bundles that mnsist of from 36 to 1~ fuel m h some assembled in a square array. zirmnium steel for to economic advantages as the same fuel bumup bemuse of its lower neutron capture characteristics. The selection of clad materiall pellet diameter, and other details of the fuel assembly depends upon the design optimization for the particular application. There is a significant amountof experience with metallic fuel typesfrom the naval remtor program. by a high enrichment of These fuels are uraium-235 and are usually fabricated in the form of a msembled into a single fuel multitude of fiat a so-called ~~~h fuel plate is composed assembly. picture frame construction where the uranium metal is a sandwich with ,,ladding material on each edge and on both front and back surfaees. The uranium metal is usually metallurgcally bonded to the fuel cladding to improve heat transfer. ciharacteristiCs generally attributed to metal fuels are: (1) high heavymatomdemity; (2) a significant and reliable thermal-expansion coe~cient;(3) amenability to potentially inexpensive fabrication rnetho&j; and (4) high thermal conductivity. ~~~~~~ldisadvantages of metal (!) low melting temperatures; (2) high rates of fuels radiation-induced swelling; and (3) poor high-temperature compatibility with austenitio stainless steels. A high thermal conductivity and low melting temperature tend to ofisat each other in terms of the specific power attainable, but metal fuels have the potential for somewhat higher specific powers than oxides. may be considered. ~h~~~are other typesof fuels

conventional plant where the maximum temperatuE in limited by the chemical reaction of fuel oxidation and rate of energy release is a direct function Of the rate of fuel injection, a nuclear reactor has no such limitation A nuclear reactor hm a large quantity of in the fuel contained within the Emtor) and the maximum temperature of the reaction is limited only by th'! ability to remove heat or, more properly) by the by spondence between the heat removed fmm the as a functio'l the coolant and the Power level of the ,f the excess reactivity or neutl'0niCs Of the systemThis should be recognized as being true only On a interest for retical basis, since for Power reactors marine pr0pUlSion the neutr0nics Of the system are 'I' power that operation at Power levels above removal of the moderator, which has a negative Or "shutdown" effect on the mactor, and all major 'ystemH are designed to fail safe or shut the reactor dew''' Nevertheless, the point is still valid that generally 'I1'' removl'l most important aspect of l'HiCtor design is Of heat and the most important single the coolant selection. A number Of possible gas molants for reactor systen'* have been considered. However, most Of the pOn*iOr bilities can be eliminated, either by 'Ir lurgical evaluations (air, hydrogen, carbon by heat-transfer considerations (neon! argon). n''' ties of gases which are suitable for reactor shown in Table [lo]. I n addition to relatively Poor heat transfer,am)1' and neon also has problems of neutrofl very expensive. N i t w e n has a high n'utrOn-absorptiO1' cross section and might cause nitriding at high tures. Thus, the list of gaseous Coolants of interest fol marine pmpulsion can be reduced to carbon dioxi(it"

''

141

Properties of Gases Suitable for Reactor H2 2

He 4

0.125 0.199

...

0.097 0.135 0.172

0.010 0.015 0.020

0.023 0.033 0.044

3.47 3.51 3.60 0.0052

1.24 1.24 1.24 0.0104

0.249 0.259 0.279 0.0727

0 0178

0.0129

1.19 0.17

Na

cooling

Air 29

CO 28

0.018 0.028

COr 44

A 40

0.017 0.027

0.013 0.042 0.028

0.012 0.025 0.018

0.020 0.031 0.044

0.017 0.041 0.028

0.027 0.054 0.041

0.241 0.254 0.272 0.0748

0.250 0.262 0.283 0.0727

0.217 0.262 0.295 0.114

0.124 0.124 0.124 0.104

0.0180

0.0179

0.0180

0.0238

0.0129

1 .OO

0.73

0.73

0.72

0.79

0.68

1 .oo

2.2

2.2

2.2

4.0 10 9294 0.0456

4 0 0 7225 1284

4.0 60 0.51 0.5

0.17 1 .O 6 22.7 4.53 X 10-4 18.5 0 0

28 0.018 0.028 0.037

0.020 0.031 0.041

0.039 0.021 0.032 0.042

...

\

0.88

10

1.8 5 1 .O 1 .O

24

40 1392 137,065

(excess N?aCtiVityto overcome the poison effect Of shortlived radioisotopes immediately after shutdown) Since the excess fuel a t start-up provides reactivity in an excess of that required to maintain essential aspect of reactor control is to provide margin There are, however, a number of disadvantages of for shutdown at all conditions. In addition, since the water as a reactor coolant. As more advanced tech- power output of a given reactor is directly pmportional nology is developed, it is probable that water will be to the neutron density or the number of neutrons Per replaced by a reactor coolant that will permit more unit volume fuel, the control system must sense and reactors. The general limitations associated limit any excessive rise of neutron flux during power level with the use of water as a reactor Coolant are:

A water coolant provides the capability of direct steam generation in a boiling-water reactor. Water technologY is well known and system cornponents are available, reliable, and relatively inexpensive.

hecame, although the fast neutrons are slowed down to herma1 energy, there is excessive neutron absorption in tho water as compared to fissile capture of neutrons in NEUTRON

-

ABSORPTION Several important control characteristics of lightI N MODERATOR water reactors Can be observed from Fig. 3. The most l'nportant is that light-water reactors are nonauto(jlltal~ticin that, if the reactor power is increased (even transiently) above the ability of the cooling system to IVmove heat, the moderator-to-fuel ratio is reduced, a I ~'roviding a negative reactivity or shutdown effect. Itemoval of moderator from the fuel region may be llccom~lishedby either steam void formation or by njoction of water. I n the case of boiling-water reactors which are normally designed to operate slightly underWATER-TO-FUEL RATIO and be provided Fig. 3 Variation in reactiGv as a function of wo+er-to~fuelratio for an '10take care of reactivity lost due to steam voids. From idealized, homogeneous, thermally critical lightmwater Ipig. 3 it can be seen that for undermoderated systems a #hamvoid would displace some of the moderator, ""'ulting in a 'light reduction of reactivity. I n addition, good moderators, they the energies of very fast in temperature (mide from spectral effects) and neutmns ss a result of inelasticacattering collisions~ i'mssure a moderator (and therefore Elements such as lead, barium, or imn readily decrease a)o1ant) can be expectedtoresult in changes in reactivity. the neutron energy down to about 0.05 M~~ where the r'ight-water are designed to have a negative hydmgen (elmtic) scattering cross is relatively moderator temperature coefficient. Themfore, a Cold large. Hence a combination of a moderately heavy or r"ctor that is but has not reached operating heavy element with hydrogen will slow down lamperature will be subcritical a t operating temperature. even neutmns of very high energies. this provides good operating characteristics, Essentially, every neutron that undergoes an inelastic da~ending 'POn the magnitude of reactivity swing collision is because of tfie high between hot and it does require sufficient excess probability subsequent slowing down and capture. rwctivity to shut down in the cold condition. Further, even in an elastic collision, in which case the 2'5 Shielding' For such as marine decrease in energy may not be large, the acmmpanying m"ctors, "IMiderable design attention must be given to change in the direction of motion of the leads to the attenuation of emitted nuclear radiations by an i n c m e d length of path through the shield such that lome Of Not only is such shielding the probability of slowing down and capture ia thereby n@oessaq for the protection of Personnel, but a high increased. Consequently, as a first approximation the r"diation backmund will interfere with the operation effectivenWsof a particular material for the attenuation of used in various aspects of reactor Opera- of fast neutrons is determined by the total fast-neutron cross section, which includes both inelastic and elastic the radiation a reactor System includes scattering as well as direct capture. 'Ipha and beta particles, gamma rays, and neutrons Of For maximum efficiency,a shield should attenuatefast 'brious energies, Only gamma rays and neutrons need be neutrons and gamma rays at such a rate that their fluxes since these are far the most penetrating. will be reduced to the maximum permissible values at Any which attenuates these radiations to a the exterior of the shield. hi^ requirement be "lfiCient reduce all the others met if a of high mavl number and hydrogen (or b negligible value. hydrogen pompound) were uniformly distributed in the the reactor three aspects; namely, proper proportions throughout the shield; hbwever, 'lowing down the fast neutrons, capturing the slowed- this is generally not a possibility for shipboard shielding down and forms of gamma both because of ship arrangements and also because of ruliution. since of low mass numbers are the structural requirements for the heavy used. bat moderators, hydrogen in the form of water can In marine propulsion maohm, the weight of shield @'Itably be used ss the shield constituent for slowing is of major importance; if the shielding is too heavy, the fast neutrons' However, at high neutron energies reactor may not be suitable for its intended purpose. 'Iis acattenng cross Of h ~ d m g e nis Very small; this instance, the cost of the shield may be considerable thicknm of hydrogeneous secondary in significance. In addition, shielding that lluterial be required to down the fission results in a relatively concentrated loading distribution '@'limns Of highest energy. The situation can be must be carefully considered since such load distributions impmved by an element of fairly may lead to problems with the ship,s structure. 'lKll mass such substances are not Wherever possible, advantage is taken in shield design

NUCLEAR MARINE PROPULSION

MARINE ENGINEERING

of the attenuating effect of distance, according to the categories according to their functions: (1) heavy elements to absorb the gamma radiation and slow down inverse-square law, on the radiation intensity or flux. very fast neutrons to about 0.05 Mev by inelastic colliIf the operating personnel can be kept a t an appreciable sions; (2) hydrogenous substances to moderate neutrons distance while the reactor is in operation, a significant saving in thickness of the shielding may be feasible. For having energies in the range below about 0.05 Mev; and example, a shield may be made thinner at the top and (3) materials, notably those containing boron, which capture neutrons without producing high-energy gamma bottom if access is restricted to the sides. To protect the heavy structural components surround- rays. Heavy elements which have been employed in metallic ing the core from possible damage from the heat form for ship shielding are iron and lead. Iron turnings liberated upon absorption of radiation, a so-called thermal shield is frequently introduced close to the or punchings, as well as iron oxides, have been incorreactor. It consists of a substantial thickness of a porated in concrete for shielding purposes. Because of its high density and ease of fabrication, dense metal of fairly high melting point (e.g., iron) lead is a good shield component. For gamma rays with placed between the reactor core and the main shield, or biological shield (see Fig. 1). The thermal shield energies in the region of 2 Mev, roughly the same mass lead as of iron is required to absorb a specified fraction consists of a material which effectively absorbs gamma of radiation and inelastically scatters fast neutrons. Since of the radiation. However, a t both higher and lower these two types of radiation carry most of the energy energies, the mass absorption efficiency of lead ill leaking from the reactor, a large amount of the heat appreciably greater than that of iron. The disadvantages of lead in reactor shields are its ~roducedin the shield will be released in the thermal relatively low melting point and its softness. It cannot shield. carry any appreciable portion of the reactor system If the circumstances are such that passengers or other load and, because of relatively low temperature limits, it ship's personnel can be kept at a good distance from a may require cooling. reactor when it is in operation, it is usually desirable to Masonite, with a density of about 1.3 g/cu cm, wun do so. This may be accomplished by designating used as the hydrogenous material in some of the early exclusion areas of several maximum permissible radiation reactors. The number of hydrogen atoms per cubio levels for passengers, ship's crew, and reactor operators centimeter is not much less than that for water. Is on watch. The reflector makes an important contribution to fast- addition it contains both carbon and oxygen, which can neutron shielding. The reflector, especially for a thermal act as moderators. As a general shield material, there is much to recomreactor, is invariably a good moderator (e.g., water, mend concrete since it is strong, inexpensive, tllltl heavy water, beryllium, beryllium oxide, or graphite) adaptable both block and monolithic types of tollso that it will slow down an appreciable portion of the struction. toOrdinary concrete of 2.3 g/cu cm density moderately fast neutrons escaping from the core. contains somewhat less than 10 percent by weight of Because of scattering, many of these slowed-down water when cured. Although the hydrogen concentrrbneutrons are returned to the core, thereby easing the tion in concrete is considerably less than the concenh~lshielding problem. tion in water, the larger proportion of oxygen (whioh ---An imoortant function of reactor shielding is to acts as an additional moderator) and the calcium ntlrl capture the neutrons after they have been slowed down. silicon in concrete compensate, to a great extent, for thcr This is done by inelastic scattering and subsequent difference. Nevertheless, ordinary concrete alone is nol capture by materials in the shield that have a large very efficient a s a reactor shield material since it normtd ly neutron capture probability. This is accomplished if a contains no element of high mass number. good moderating element such as hydrogen is present in Various special ("heavy ") concretes incorporati IIK addition to materials of medium or high mass number. heavy elements have been developed for reactor shieldi~~y. I n addition, an effective shield provides for the absorpIn barytes concrete, for example, the mineral b a r y h ~ ~ , tion of the various primary and secondary gamma rays. consisting mainly of barium sulfate, largely replaces t h ~ The penetration of gamma rays is a function of their sand and gravel aggregate in ordinary concrete. 'I'l~n energy but they are effectively absorbed by a material density of barytes concrete is about 3.5 g/cu cm. Thw ti of high density. The shield material, such as iron or shield of barytes concrete would have to be no thiokrr lead, which serves as the inelastic scatterer of neutrons than an iron-water shield of the same effectivcncn~ will also function as the absorber of gamma radiation. although the total weight of the barytes concrete shioltl Within the energy range of interest, gamma absorption would be greater. is determined essentially by the mass of the shielding 2.6 Safety [14]. Nuclear ships must comply wiC11 material. The thickness of shield required to produce a the rules and regulations of the cognizant agencierr, specified absorption of gamma rays is inversely proportional to the density of the shielding material. Thus a inc1uding:athe United States Atomic Energy Comrni~~iol~ smaller volume of lead than of iron would be required, [15-191; United States Coast Guard [20-231; Uniloll States Department of Commerce; National Bureau of but the masses would be approximately the same. Shield materials may be divided into three broad Standards [24]; International Convention for the Snfnbr - --A

life at Sea [25]; the classification societies [26]; and I411t: rules and regulations of agencies having cognizance over the ports of call [27]. Attention is called to indusi8t+ir~l safety codes, which may be applicable in part. ( lodes of this type include the American Society of Mtwhanical Engineers Boiler and Pressure Vessel Code JYNJ, and the applicable ANSI and IEEE codes )2!),301. A riuclear ship should also provide a degree of safety foruthe non-nuclear portions sufficiently high to ensure rrdo operation of the entire ship. I n this respect proviriolls such as watertight subdivision, stability, fire protection, bilge pumping, fire extinguishing, electrical I~intullations,steering gear, astern power, and navigal,iotlul aids should be evaluated in order to provide for tjl~omaximum practicable safety for the ship. a. Containment. Containment constitutes the outer rrl~c:losureor other systems or arrangements which are provided to prevent the uncontrolled release of hazardous atnounts of radioactivity to normallv accessible snaceu --or t,he ship's environment in the of an accident or ,tl,\lfunction of the nuclear system. I t is tlInt any one of several containment methods may be ,no& suitable for a particular application. Separate prossure-tight containment vessels or containments ~~bilifiing integral portions of the ship's structure are rrtttnples of containment systems that may be utilized. 111 the design of a containment system, the effects of pul-ification of radioactive loops, pressure relief or luppression systems,and systemswhich effectively pmvont core meltdown or its consequences, should be oI'

event

r -

145

should, therefore, be designed to contain, control, and possibly suppress the release of radioactive material which could result from any credible accident. Consideration should be given to (1) the pressure and temperature of the coolant, (2) the energy released as a result of any chemical reaction within the system, (3) the nuclear heat generation, including afterheat, and (4) the energy stored in the structure. The processes involved in the release of this energy are heavily dependent upon the type and specific design of the nuclear power plant. Each system should be evaluated on an individual basis to determine the pressure buildup in relation to the containment d-esign. Missiles resulting from a malfunction of the system components should not result in the release'of hazardous amounts of radioactive or toxic materials to occupied spaces or the ship's environment. The following components are typical of those which may be considered as potential sources of missiles: High-speed rotating equipment. The installation such withinthe should be kept to a minimum, but, if installed within the containment~ be to reduce the probability of rupture of the containment wall due to a failure of any "tating Rods. Positive means be provided to prevent rods from being ejected. ' within the pressurized system. These should be located or protected so as to minimize the p"bability damage to the containment walls in case of failure.

'he containment system should be designed to ensure The primary objective under these circumstances the basic integrity of the containment will be should be to maintain the integrity of the containment tained for any credible operating or twcident and, insofar as practicable, to prevent impairment of the The following factors are typical of those secondary shielding when materials particularly sushiah should be considered: ceptible to fire damage are used (e.g., lead, polyethylene, Or r Maximum credible pressure buildup within the The containment should be designed to remain intact b~tninmentdue to an accident to the nuclear system. if the ship sinks in shallow water, and consideration r Maximum credible internal missile. should be given to provisions for decay heat removal. 0 Location as regards collision or grounding damage. r Itupture of piping, ducts, or similar components Containment integrity should be maintained for a ~ i d eof the containment, and such components con- period of several years following such an incident in order to provide sufficient time for salvage operations. tod to and passing t h r o ~ g h ~ t hcontainment. e b. Shielding and Radiological Safety. Shipboard r External fires and explosions on board. shielding and radiological safety are intended to provide Fires within the containment. standardd for protection against nuclear radiation for 0 Binking of the ship. personnel on board ship and for persons in the vicinity Forces due to ship motion. of such ships in conformance with the cognizant regular Itemoval of reactor decay heat in the event of loss 81 aeolant circulation and provisions for preventing the tory agencies. Inasmuch as all regulatory agencies normally follow the recommendations of the Federal !@&atorcore from melting through the containment. Radiation Council [16, 171, the recommendations of the e Leakage and measurement of leakage rate. Federal Radiation Council should be considered to All nuclear systems producing useful power contain anticipate changes to the criteria specified by the regula@ ~ O T Oenergy ~ indicated by pressure and temperature. tory agencies. It is the intent to provide standards for l ~ d d r nuncontrolled release of this energy and any protection by means of shielding and control of personnel &idltiunal energy that might be generated in a nuclear access so that passengers and shore personnel will not be ~ ~ l d ( !provides nt a potential mechanism for the diaper- exposed to radiation exceeding recommendations for the ~lo11of radioactive material. The containment system general population, and so that operating, maintenance,

.

MARINE ENGINEERING Limits for Table Liquid waste Disporal Discharge to h e Sea as Specified for the NS Savannah

Table 3

NUCLEAR MARINE PROPULSION Radioactive GasWane Specified for the NS Savannah LIMITSON

limb as

a~aseouswaste discharges are to be made while the underway.

is

from shoreline at

depths greater than 200 fathomsd

be used to pmtect against missiles and to provide con[311' The prime function "I d- Health tainment in case of an accident to reactor co~pOneIltS. individu'd' waste DiSposal. Radioactive wastes health physics is to safeguard the to nucl"lL' c. resulting from the hSion whose work is likely to are defined rn the end which contain radioisotopes in significant radiations by taking all steps that are 'Onsidered ncc(u liquids, sary to minimize such exPosure. In addition there "11' quantities. Radioactive wastes include and grnes. Some examples of solid wastes are con- responsibility of making sure that nothing escaping fr''ttl p ~ i C land ~ ,spent the nuclear plant, even in the event Of an accid('l'l" taminated dirt, ,,hips, or other a ~ ~ ~ which i has ~ become ~ e Would ~ represent t ion exchange the ma*mum ranit'' The regulatory bodies contaminated or radioactive may also have to be treated tion exposure limits for personnel, maximum Permis~il''' in The the same manner a waste. purpose of a radiosetive wastedisposal system is concentrations of certain radioisotopes in air and wd''l' and dispose of waste material to mlleot, audit, of any area in a manner that limits the

and maximum permissible amounts Of such Such may accumulate in the human body.

"lN'

rcc('lll

"'cl'datiom are subject to regular review a. increasing ""owledge is g ~ n e Of d the effects of nuclear radiation on '"lf'''lrt ' human body' Dosages are set at such low levels Over many years is unlikely to cause injury. On the Other hand, the levels cannot be so low as to make operation of a plant impossible. 011 board One of the Primary r~pomibilitiesof I'O"lthphysicsisto monitor radiation. This involves the 'l('tcrmination and recording of radiation dosages and (lt)~c rates at nUmerOUS locations. Radiation dosage is rrloasured in terms of the energy absorbed from the radiaIli(l11~ and the dose rate is the time rate at which such f'Jlorg~ is absorbed. I n general, the total dose (or dosage) ~*nouived is the product of the dose rate and the exposure

curve observed on large central-station nuclear plants has not been apparent in unitsof lower power levels. I n fact, in 1969 the product lines for several manufacturer's of central-station units did not include power ratings as low as 300 megawatts electric lmw(e)l. ThereforeJ the Capital costs of around $2@)-$220/kw(e) for to capital costs of grestermw(e) than units must be for unitsthat, of 50based to onmw(e). In addition, it may kw(e) be expected parisom of fossil-fired marine and stationary units, mobile Power Plants will cost about 35 percent more than land-based units of the same rating. Studies 133-351 have indicated that fol large marine reactors of 70,000 to 50,000 shp, fuel costs will be as low as 2.2-2.0 mill/shp-

NUCLEAR MARINE PROPULSION

1A8

Table 4 Summary of Fuel Cost Data P L A N A T CORE Basis-year Natural U $/lb UIO~ Conversion $/kgU Separative work $/unit Tmls com osition % (U-235g f t after . separation processing) Pu credit $/gm fissile Fabrication $ / I t @ Spent fuel shlpplng Reprocessing, $/kgU Reconversion, $/kgU Capacity factor

MARINE 1 2 1977 1974 8.55 8.10 2.29 2.29 26.00 20.00 0.20 0.20

Nf3

Swannah

3 1981 8.80 2.29 20.00 0.20

1 1968 8.00 2.50 30.00 2.53

8.55 7.89 8.00 114.00 100.50 87.50 6.00 6.00 6.00 29.10 31.80 31.80 3.00 3.00 3.00 70% 70% 70%

10.00 88.75 6.00 52.35 5.60

.. .

Table 5 Nuclear Fuel Costs, mills shp-hr PROPOSED PLANT A

NS

Savannah

CORE

1 1974

2 1977

3 1981

1

Resis 1968 - -Direct costs 1.107 0.957 0.934 1.777 net uranium -0.246 -0.200 -0.194 -0.450 plutonium credit 0.611 0.426 0.335 1.667 fabrication shipping, repro., recon. 0.204 0.173 0.156 1.201 subtotal direct costs 1.676 1.356 1.231 4.195 Working capital Outof-m-


The displacement is then multiplied by the appropriate ratio of engine weight to displacement (generally 4) to determine the engine weight. The space requirements of a diesel power plant are rather flexible in that it is possible to assemble a plant from one or more units and to select the type of unit to be used. If head room is a ~roblem,small high-speed engines can be used. If width is a ~roblem,in-line engines can be used. If it is necessary to minimize the length, vee-type engi~lesare available. ~ n g i n e scan be furnished completely assembled with all the necessary accessories mounted on the engine and its subbase, or with these accessories loose for mounting where space is available. It is extremely important that adequate t space be ~rovidedaround each engine to ~ e r m i access for maintenance. Fortunately, the space required for maintenance usually coincides with the envelope of the engine. Parts of high-speed engines are relatively small and light in weight; this facilitates handling and minimizes the need for extensive rigging for art removal.

3.2 Shipboard Applications of Diesels. Diesel engines are used either singly or in multiple to drive propeller shafts. For all but high-speed boats, the modern American diesel turns too fast to drive the propeller directly with good efficiency and some means of speed reduction, either mechanical or electrical, is necessary. If a single engine of the power required for a given application is available, then a decision must be made as to whether it or several smaller engines should be used. This decision may be dictated by the available space. Using a mechanical transmission system as an example, Fig. 11 illustrates the flexibility of the diesel power plant in adapting to specific space requirements. I n this figure, an engine with a rating of 5000 bhp irs used

\

In applications where it is necessary to provide rapid maneuvering characteristics with reverse gears or direct reversing bngines, brakes may be installed either on the propeller shaft or on the high-speed pinion shafts of the reduction gear to stop the propeller shaft in minimum time. Many direct reversing engines can be specially adapted to use starting air in the cylinders for braking purposes, and this possibility should be weighed against other means of shaft stopping. Diesel engines are used to drive shipsJ and emergency generators. Emergency generator sets are arranged to start automatically upon failure of the normal power supply, and after a builtrin time delay, assume the electrical load on the emergency bus. For many years, Navy specifications have required that

,I

I 1

I

I

MEDIUM AND HIGH-SPEED DIESEL ENGINES

MARINE ENGINEERING REDUCTION GEAR OUTPUT 170 RPM

I I V-16 ENGINE ( 5 0 0 0 BHP AT 514 RPM) HEIGHT 10'

REFERRED CURVE

2 V - 8 ENGINES (EACH 2 5 0 0 BHP AT 514 RPM) HEIGHT 9.5'

IT (0) .2V-16 ENGINES (EACH 2500,BHP HEIGHT 7.7

20

30

40

50 60 70 80 ENGINE OR PROPELLER RPM.% RATED

90

100

110

Fig. 12 Matching engine to l i p characteristicswing power cunes

I UI

I

2 V-16 ENGINES (EACH 2 5 0 0 B H P AT 7 3 0 RPM) H E I G H T 7.7'

38.5'

m 4 V-16 ENGINES (EACH 1250 BHPAT 1 0 3 0 RPM) HEIGHT 5'

-1

2564 .-'

Fig. 1 1

Compar'wn of various engine arrangements for 5000-bhp plant

emergency generator sets be capable of starting and assuming full load in no more than ten seconds, and it has been demonstrated that this is a reasonable require ment. It is possible to parallel a diesel generator set electrically with generators driven by other diesels or other prime movers such as steam or gas turbines; however, the equipment supplier should be made aware of this requirement if it is needed. Diesels are used as prime movers to power many different types of auxiliaries such as fire pumps, dewatering pumps, cargo oil pumps, compressors, and winches. For engines installed high in the ship, conideration should be given to the use of radiator-cooled or air-cooled engines to avoid cooling water pumps which would be subjected to high suction l i t requirements. 3.3 Selection of Engines. The selection of engines for shipboard use cannot be b-d on any single factor.

There are many possible engine deaigns which are capable of meeting most performance requirements, and numerous factors must be considered such as weight, fuel consumption, cost, availability of competitive engines, manning requirements in terms of skill level and number, and maintenance considerations such as availability of repair parts, necessity for special tools, and the number, type, and frequency of the maintenance required. References [17-221 discuss this subject in detail. The first and possibly the most important consideration leading to the selection of a diesel engine is the definition of what it must do. I n the case of a propulsion engine, this entails obtaining the speed-power curves for d l important modes of operation such as fully and lightly loaded, clean and fouled bottom, towing and running free, and with and without power takeoff loads. Additional information should be aster-

tained regarding the time duration of operations a t each condition. An assessment should also be made of anticipated special operating requirements. For exill it be necessary to spend long periods of time ample: W with engines idling? Will long periods of slow-speed maneuvering be required? Will the operation be primarily point to point with the engines a t full load and speed most of the time? Each of these questions and many more can influence the design of the diesel power plant. When the speed-power curve has been established, an engine can be selected which will develop the required horsepower a t its appropriate rating. Assuming that the ship under considerati~nis one which is expected to operate the majority of its time a t less than full load, the intermittent duty rating would be the appropriate one. A particular engine, or engines, is then selected whose intermittent rating is consistent with the full-power requirements for the ship. The intermittent horse power curve for the engine, similar to Fig. 10, and the light-load lines from Fig. 4 are then superimposed on the speed-power curve. Preferred, acceptable, and lightload operating regions are then added and the resultant plot is illustrated by Fig. 12. Operation in the light-load region should be avoided. The propeller load, curve A in Fig. 12, has been drawn with the power varying as the cube of the s p e d . It can b e seen that operation down to about 70 percent speed is within the preferred zone, and from 70 to 55 percent speed is in the acceptable zone. If ap appreciable amount of time is to be spent in operation below 55 percent speed, where the engine load as dictated by the speed-power curve faIls into the undesired rarige, consideration should be given to the use of two or more engines instead of one. Curve A represents the power to drive the ship with a clean bottom whether that power is produced by one engine or multiple engines. If

the performance of one of two installed engines operating alone is to be evaluated, it is necessary to redraw either the engine performance curves or the speed-power curve. Either the engine performance curves would have to be drawn with ordinates one half their original magnitude or the speed-power curve would have to be drawn with ordinates twice its original magnitude. It is simpler to redraw the speed-power curve, and this is shown as curve B. Now it can be seen t h ~ one t engine can be declutched from the propeller shaft whenever the ship speed is reduced to 62 percent of full speed (the intersection of curve B and the continuous-duty line). Under these conditions, the single engine would operate in the recommended zone, whereas two-engine operation would be in the acceptable zone. At speeds down to about 38 percent, the single engine would be acceptable, whereas two engines would be too lightly loaded below 55Ifpercent low-speed speed.operation is required for substantial periods of time, consideration should be given to a larger number of engines. Using the same procedure as previously, curve C has been drawn to represent the speed-power curve when operating on one fourth of the installed engines. I n this case, one engine could be used for operations up to about 40 percent speed, two engines from about 40 to 62 percent speed, three engines from 62 to 75 percent speed and all four engines above that. I n addition to the improved loading condition of the engines during part-load operations, benefits are derived from the fact that only some of the engines accumulate operating hours, and the total fuel consump tion is less. I n c w s where the speed-power curve can vary with conditions of operation (e.g., different displacements, water depth, hu\l fouling, towing), the extremes of loading should be considered when selecting the engine-

260

MARINE-ENGINEERING

MEDIUM AND HIGH-SPEED DIESEL ENGINES

have a power rating less than 10 percent of that of the main engine and be disengaged when the main engine is used. The quick starting capability of the diesel obviates the necessity of keeping engines running at idle just so that they will be ready when needed. The characteristics of diesel engines and the principles

26 1

governing their proper selection and application have deliberately been expressed in general terms. By following the methods described, unusual applications such as the engine requirements of planing hull boats or hydrofoils can be handled as well as the more conventional ships and boats.

Section 4 Design Considmtions

0

20

40

60

80

100

ENGINE RPM.% RATED Fig. 13 Matching engine to ship characterlrtiu uaing torque CUNW

propeller-reduction gear combination. Curves A' and A" which represent these extremes have been added to Fig. 12 to show the effect on performance. If the ship were designed to absorb full power under the conditions of curve A and then were required to tow a load such that the total resistance corresponded to curve A', the maximum speed permissible would be 85 percent of rated (the intersection of curve A' and the intermittent rating curve); the limiting factor would be engine torque. If, on the other hand, the resistance were reduced to that shown by curve A", no speed increase would be possible without overspeeding the engine, and full engine power could not be utilized. Under these conditions, the choice is dictated by the condition under which it is most important that full power be developed. If full power is required under both conditions, a controllablepitch propeller or a two-speed reduction gear must be used. Figure 13 shows the same conditions plotted with torque and rpm as coordinates to illustrate an alternative method which could be advantageous when most data are available in that form. The engine torque curves

shown-in Fig. 13 are not consistent with the horsepower curves in Fig. 12, which were drawn as straight lines for simplicity. In addition, specific fuel consumption curves have been added. The reduction in fuel consumption at low speeds which is obtained by operation with reduced numbers of engines may be verified from this plot. Figure 13 can also be used to verify that the minimum specific fuel consumption of the selected engine occurs at the ship speed and load most frequently expected. It is not necessary that all engines in a multi-engine drive be identical, although logistics problems are simplified if they are. There are cases where a considerable amount of low-speed maneuvering is required, and, if the required speed is below that corresponding to engine idling speed, the low speed can be obtained by the use of CRP propellers, two-speed transmissions, slipping clutches, or the use of a small engine which is geared to the propeller shaft such that it develops full power at a ship speed slightly above that corresponding to the idling speed of the main engine. The small engine may

4.1 Types of Fuel Used. One of the prime objectives in the development of the diesel engine has been to provide a prime mover which would be capable of burning a wide variety of fuels. It has, however, been necessary to compromise on this goal in order to achieve others such as reduced weight and space, increased reliability, lower wear of parts, good cold starting ability, and increased safety in fuel handling and storage. Over the years, a number of specifications for fuel oil have been developed to insure that the customer would be able to buy fuels meeting the requirements of various engine designs and t o give new engine designers a range of standard fuels from which to select. Operators who maintain a fleet of ships are particularly desirous of supplying one grade of fuel for all of their engines. The most significant characteristics of diesel engine fuels are listed in Table 3. The generally accepted uses for these fuels are: ASTM ID. A volatile distillate fuel oil for engines in service requiring frequent speed and load changes. The flash point of this fuel should be specified as a minimum of 140 F for marine applications. ASTM 2D. A distillate fuel oil of lower volatility for engines in industrial and heavy mobile service. Again a minimum flash point of 140 F is recommended for marine service. ASTM 4D. A fuel oil for low and medium-speed engines; however, it should not be assumed that all low and medium-speed engines will run successfully on this grade of fuel. The advice of the engine manufacturer ,should be solicited before using grade ASTM 4D fuel to insure that the particular engine model can tolerate the wider range of fuel properties permitted by this specification. MIL-F-16884, Marine Diesel Fuel. This Navy specification fuel is generally similar to ASTM 2D fuel except that a higher cetane number and flash point are specified and particular attention is paid to insure that fuels from different sources and lots will be miscible and that good st0rake stability is provided. MIL-T-5624, Turbine Fuel, Aviation Grade JP-5. This fuel is similar to ASTM 1D fuel except for its lower end point and high flash point. It has many require ments which are not tabulated in Table 3 inasmuch as they are needed primarily to meet aviation engine

reguirements, and they are not relevant for marine ap$ications. JP-5 fuel must be provided for turbinepowered aircraft o'perated from ships at sea; therefore, this fuel is used by the Navy in all diesels which are refueled at sea in order to simplify logistics problems. JP-5 can be used successfully in diesels while MIL-F16884 fuel cannot be used in turbines for aircraft use. The increased cost of the JP-5 fuel is offset by the advantage of having to carry only one grade of distillate fuel in tankers. The relationship between engine performance and some of the fuel characteristics specified in Table 3 is as follows: Cetane Number. Cetane number is a measure of the ignition quality of the fuel. Engine performance factors influenced in part by ignition quality are: (a) cold starting, (b) warmup, (c) combustion roughness, (d) deposits under idle and light-load operation, and (e) exhaust smoke density. Each of these performance factors is also affected by other fuel characteristics and engine design parameters. The cetane number requirements of an engine depend on design, size, mechanical condition, operating conditions, atmospheric temperature, and altitude. An increase in cetane number o?er values actually required does not materially improve engine performance. Heating Value. This important property of a diesel fuel is a measure of the energy available from it. The heating value of fuels may be expressed in either of two ways: high or gross heating value and low or net heating value, the difference being the latent heat of the water in the exhaust gas. Heating values may be expressed in terms of Btu/lb or Btu/gal. Since diesel engine fuel consumption is normally quoted in terms of lb/hp-hr and fuel is purchased on the basis of cost per gallon, Btu values on both a weight and volume basis are of interest. It is now customary to use the lower heating value for calculating thermal efficiency of diesels, although in the past the higher heating value was used. In either case, thermal efficiency is of academic interest only. For a comparison of the performance of different engines on different fuels, fuel consumption in terms of Btu/hp-hr is most useful, although care must be taken to insure that the heating values of the fuels are reported on the same basis. The heating value is specified in only one of the specifications listed in Table 3; this is because distillate

.

262

MEDIUM AND HIGH-SPEED DIESEL ENGINES

MARINE ENGINEERING Table 3

Ignition quality-*tam no.. ............... Appearance. .............................

Diesel Fuels ASTMD 976-64T 1D 2D 4D -40(d 40(a) 30'"

Marine Diesel MIGF 16884

JP 5 MILT 56246:

45

SEPARATOR

.. 1 1 1

.

DIESEL OIL SERVICE TANK I $DRAIN L. VLV.

IF VALVE IS USED ALSO INSTALL PRESSURE REL.VLV.

A

STRAINER

FILTER

r

Pour oint, F (mu). ...................... l (e )P.. ~ i ~ ~ ~F ~ ~ i..................... n t , 1.4 Viecoslty @ 100 F: centistokesmin. ......... m u . ........ 2.5 SSU min ............... gax .............. 34.4 0.15 Carbon residue on 10% bctttom, % m a . . ... 0.50 Sulpbur, % ( m u by we1 ht.. ............. Corrodon (mu) )at 212 Fy .;............... No. 3 Color (ma). ............................. Ash, .% (m)by weight. ............. : ... 0.01 Gmwty, API mm/max. ................... Acid number (max). ...................... Nel,ltrrtlity.. .............................. M m e pomt F. ........ !................. Accelerated stability%otal.................. insolubles, mg/100 ml (max). ............ Water and sediment b volume %. ......... TRACE Lower h e ~ t e "sale, g itu/lb (min). ......... Aniliue gramty product (min) ..............

125 or legal 2.0 6.8 32.6 45 0.35 1.0 No. 3

5.8 26.4

45

125

2.0

D.O. MANIFOLD

2.1 6.0 33

D.O. INJECTION PUMPS

45 0.2 1.0 No. 1 0.005 Record 0.50 NEUT Record 2.5

FINAL

k UNIT INJECTORS

0.4 No. 1

6

0.10

-

u

FUEL SUPPLY PUMP

(b)

Fig. 14 Dlagram of a typical fuel system

36/48

NOGS: (a) Lower tem rature or high-altitude operation may require higher cetane number. B below the minimum expected ambient. b S ecify at R r comparison not a apeyification d u e . d) For test methods, see specification referenced.

Go

diesel fuel properties such as volatility, viscosity, gravity, ignition quality, and heating value exhibit interrelationships. It has been established that certain characteristics of fuel can be estimated with reasonable accuracy from two or more measured characteristics such as volatility and API gravity. Charts ahowing these relationships may be found in the SAE Handbook [16] in the section on diesel fuels. For estimation purposes in ship design, a fuel with a representative higher heating value of 19,350 Btu/lb can be used. The corresponding lower heating value is 18,190 Btu/lb. Engine performance on the test stand is corrected to reflect the diierence in the heating value of the actual fuel and the standard value used in design. Viscosity. For some engines, it is advantageous to specify a minimum fuel viscosity because of the power loss due to injection pump and injector leakage. Maximum viscosity, on the other hand, is limited by considerations involving the engine design and size and the characteristics of the injection system. Sulphur. The effect of sulphur content on engine wear and deposits appears to vary considerably in importance and depends largely on operating conditions. It is important to maintain an engine jacket water temperature of at least 140 F to minimize the effects of sulphur in the fuel.

. WATER

PUMP

clear and brinht

Diatiition, 10% point F .................. Distillation, 90% p o w (mu). ............ 550 W i t i p n ; end point F (max) ............. Flash pomt, F (mip). ...................... 100 or

P

I

TRANSFER

Flash Point. The flash point as specified is not directly related to engine performance. It is, however, of importance in connection with legal requirements and safety precautions involved in fuel handling and storage and is normally specified to meet insurance and fire regulations. For marine use, a minimum flash point of 140 F is recommended. Pour Point. Pour point is important in connection with the lowest temperature which the fuel may reach and still be sufficiently fluid to be pumped or transferred. The pour point is generally interrelated with cetane number and volatility. Frequently, low pour pointa may be obtained only at the expense of lowering the cetane rating or increasing volatility. The pour point should not be specifled lower than required. For a more comprehensive dkussion of petroleum fuels, see Chapter 23. I n the design of a new ship, the selection of the fuel ts be used has an important bearing on the selection of engines and the detail design of the fuel handling and storage system. The fuel selection may be specified by the owner or left to the ship designer to provide greater flexibility in optimizing the total design. The selection of fuel for a given engine requires consideration of the following factors: (a) fuel price and availability, (b) maintenance considerations, (c) engine sire and design,

(d) speed and load ranges, (e) frequency of speed and load changes, and Cf) ambient conditions. 4.2 Fuel Oil System Design. The fuel injection system of a diesel engine is, in many respects, the heart of the engine. It must meter extremely small quantities of fuel, deliver the metered fuel at high pressure to the engine cylinder at exactly the correct time, in a precise spray pattern, and a t a specified time stop delivery abruptly and completely. The instantaneous pressure in the fuel nozzles can be as high as 40,000 psi a t full load in unit injectors (the fuel pump and nozzle are combined into one unit with no lengthy fuel line between). In the conventional system, the fuel pressure at full load may be as high as 15,000 psi for some engines. The duration of injection in a high-speed engine can be as short as 0.001 sec. With the high pressures involved and the precise timing requirements, it is necessary to build the injection equipment with close clearances and small tolerances. Nozzle hole sises vary upward from 0.005 in. dia, while the plunger-to-barrel diametral clearances, may be as small as 1.5 microns (0.00006 in.). I n view of these small clearances and high pressures, the most important consideration in the design and layout of the fuel oil handling and supply system for a diesel engine is to insure that clean, waterfree fuel is delivered to the engine. It is particularly important in 11mrine installations to insure that there is no salt water in the fuel at the time it gets to the injection pumps and nozzles. Saltwater-contaminated fuel has been known to erode the small holes in the fuel nozzles and cause pintle corrosion and sticking in a relatively short time, resulting in loss of power, burned pistons, high fuel consumption, and a smoky exhaust. In the typical fuel system illustrated in Fig. 14, diesel oil is transferred to a diesel oil service tank, sometimes called a day tank, after passing it through a water separating device which may be either a centrifugal purifier or a coalescing-type filter. Fuel flows from the

service tank through a strainer to a fuel supply pump which is normally attached to and driven from the engine. The fuel is discharged from the pump and flows through a filter and sometimes also through a final-stage filter before going to the fuel injection pump. It is customary for the fuel supply pump to have a capacity from three to four times that actually required by the engine. The excess fuel flows through the injection pump housing, cooling the plunger and barrel and insuring that the pump cylinder is completely filled at each stroke. The high-pressure fuel is discharged from the injection pump to the fuel nozzles in each cylinder of the engine through high-pressure tubing. Excess fuel flows through leak-off lines from the injection pump and from each fuel nozzle. The leak-off lines are manifolded to return excess fuel to the service tank, d i e charging above the fuel level and preferably against a horizontal b d e . If unit injectors are used, the pump and nodsle are combined in one assembly and there is only one leak-off point from each unit. It is preferable to have a separate return line from each engine to the service tank or tanks, with no valves in the lines. If it is necessary to install shutoff valves in the return line, a pressure relief valve should be installed to by-pass the valve and discharge to one of the service tanks in case the valve is inadvertently closed while the engine is running. If cocks can be installed to divert the flow to the proper tank with no chance of a line ever being completely blocked, the relief valve can be omitted. It is possible for the pressure to build up in a closed return line to the point of rupturing the pipe, spraying fuel into the engine room, and possibly starting a fire. Care should be taken to insure that leak-off lines have a minimum number of joints and that these joints are located so that leakage will not contaminate the engine lubricating oil. The choice between a centrifugal purifier and a c o s lescer-type water separator must be made for each

MARINE ENGINEERING

installation. The purifier can be of the self-cleaning type where the dirt and water removed from the fuel is discharged to a separate collecting tank, which requires infrequent cleaning. Purifiers are available in a variety of sizes, and one unit may be able to serve the needs of all the installed engines. The initial cost of a purifier is higher than that of a coalescer type; however, maintenance costs are lower and logistic problems are simplified. Care must be taken to follow instructions carefully and select the proper ring dams or discharge rings to suit the specific gravity of each fuel being centrifuged. The centrifuge can be equipped with its own heater and transfer pump to make it capable of handling heavier fuels. A coalescer has the advantage of being a static device with no moving parts to wear out, but it does have cartridges which require replacement. It cannot be used, however, with residual fuels or distillates contaminated with residuals. I n a coalescer, a combination filter and water separator unit is used. The oil with entrained water first passes through a phenol-impregnated paper filter element where solid contaminants are rempved, and the finely dispersed water droplets are induced to conglomerate a t an accelerated rate by intimate contact through the capillary openings in the filter paper. Some of these larger water droplets fall by gravity into the water collection sump along with dirt particles. The filtered and coalesced fluid then passes on to the separator unit. Its vertically pleated element of controlled porosity is impregnated with a hydrophobic material, such as molybdenum sulphide or paraffin for preferential wetting by the oily fluid, so that the oil and not the water globules passes through the capillaries. Water is collected in the sump of the separator unit and clean, waterfree fuel is taken off from a connection &,the top of the unit. If care is not taken to change elements when the pressure drop across them exceeds the recommended limit, there is the danger of rupturing the elements and contaminating the fuel in the service tank. The strainer has a metallic element of woven wire, stacked metallic disks, or sintered metal. Woven wire elements can remove particles down to about 40 microns, and if the joints in the wire are welded they can remove particles 2 microns in size. Stacked disks are capable of removing 40-micron particles and have the advantage that they may be made self-cleaning by rotating alternate disks. Sintered metal elements can remove particles in the range of 3 to 25 microns, depending on their density. Sintered metal elements &re difficult to clean and may disintegrate if subjected to'large pressure surges. The fuel supply pump draws fuel from the diesel oil service tank through the filter, and for that reason it must have the capability of operating with a suction lift of from 4 to 6 ft. If the suction lift is too great due to the elevation of the pump or the length of the supply line or the pressure drop in the filter, a separate motor-driven fuel booster pump may be required. The fuel supply pump is of the positive displacement type with pumping elements using either gears, vanes, plungers, or dia-

phragms. These pumps \\-ill have a discharge pressure of 6 to 20 psi for small engines and 25 to 40 psi for large engines. A pressure relief valve should be provided on the discharge side of the pump, either built illto the pump housing or installed separately in the discharge pipe. The fuel from the relief valve should return to the pump suction or to the service tank. The diesel oil service tank is normally located a t a level above that of the supply pump so that fuel can be supplied to the pump by gravity. In some engines, the fuel system is so designed that the fuel service tank must be located below the supply pump to prevent the flow of fuel by gravity into the cylinders of a shutdown engine. Air leakage into the fuel inlet lines can be very troublesome; therefore a minimum number of fittings should be used and all joints must be completely airtight. This is particularly important when the fuel tank is lower than the supply pump. Diesel oil tanks should not be made of galvanized steel because of the danger of forming corrosive zinc compounds in the fuel. Copper or silicon bronze should not be used for fuel tanks either, as their reaction with the mercaptan sulphur compounds in the fuel can result in the formation of damaging copper deposits in the engine combustion chambers. Aluminum bronze and manganese bronze are satisfactory for fuel tanks, as their use does not lead to these problems. Filtration is accomplished upon discharge from the supply pump in filters containing one or more elements made of either treated paper, felt, or woven yarn. The paper elements can be expected to filter particles in the range of 3 to 5 microns, with an initial pressure drop of from 0.5 to 2.0 psi and a pressure drop of between 15 and 30 psi a t the time of replacement. Woven yarn filter elements have a greater capacity to handle dirt, higher flow capacity, and somewhat coarser particle removal capacity. It should be noted that the characteristics of filter elements of any type can vary considerably depending on the filter design. Considerations with paper filters are the porosity of the paper and the material with which it is impregnated, and in the case of woven yarn filters, the tightness of the weave and the depth of the flow path. The particle removal characteristics of a filter should be expressed in terms of particle size and the probability that that size particle will be removed; for example, 2 microns 92 percent, 5 microns 95 percent, greater than 5 microns 99.5 percent. For most diesel engines, a progressive filtering system is used consisting of filters of increasingly fine filtering ability. First there is a strainer to take out large particles, then a yarn type filter to take out particles in the 25-micron range, and lastly, a final-stage filter of the impregnated paper type to remove the finest particles. The yarn-type filter is sometimes eliminated where a clean fuel supply can be assured. In engines with unit injectors, the first-stage filters are of the paper type, and final-stage filters of a metallic type are installed in the body of each injector, one a t the inlet and one a t the outlet connections. Filters may be of simplex or duplex construction, with the latter being used when it is

MEDIUM AND HIGH-SPEED DIESEL ENGINES

not possible to shut down the engine to change filters. Again, it must be emphasized that the major objectives of the fuel system are to deliver clean fuel, free of air and water, to the injection pumps. To this end, filter cases should be installed in locations where they can be easily serviced and the elements can be replaced without introducing dirt and with a minimum of maintenance effort. Jobs that are difficult to accomplish tend to be accomplished less frequently. Steps should be taken to ensure that there is a minimum possibility of air entering the system through joints in the piping on the suction side of pumps. Adequate and easily accessible drain connections should be provided a t the lowest part of the fuel service tanks for stripping water or foreign matter which may accumulate. Systems suitable for handling heavy distillate or residual fuels are described in Chapter VIII. 4.3 Types of Lubricating Oils Used. The engine manufacturer furnishes information regarding the design as well as installation of the lubricating oil system. The manufacturer will furnish all necessary accessories and components and recommend the kind of lubricating oil to be used. Nevertheless, marine engineers should be knowledgeable of lubricating oil systems so that preliminary designs can be prepared prior to the selection of a particular engine, and to alert the engine manufacturer to unusual conditions in specific applications to insure that optimum solutions are obtained when compromises are necessary. Lubricating oils are classified into two broad categories; first by viscosity and second by the severity of the operating conditions which they can tolerate. The most common viscosity designation is by SAE numbers as shown in Table 4. Table 4 SAE Viscosity no.

5W 1OW 20W 20

30

40 60

Viscosity Values of Crankcase Oils

,

Viscosity Range Saybolt Seconds Universal at 210 F at 0 F min max min max 6,000 6,000a less than

12,000b

12,000 48,000

45 58 70 85

less than 58 70 85

110

a Minimum viscosity at 0 F may be waived provided the viscosity at 210 F is not below ... 40 - - STTS. - - -. Minim& &scosity at 0 F may be waived provided the viscosity at 210 F is not below 45 SUS.

Medium and high-speed diesel engines normally use SAE 30 or 40 lubricating oils. For small boat applications where engines are stored outdoors in cold weather, it will be necessary to use winter grades such as 5W or 10W,oils. I n addition to the viscosity, oils are classified by'the viscosity index (VI), which is representative of the slope of the viscosity-temperature curve for each oil. A high VI oil is one in which its viscosity varies little with

265

the temperature, whereas in lower VI oils the viscosity variation with temperature is greater. For engines operating in heated engine rooms, the VI is of lesser importance than in the case of exposed engines which must operate in winter a t low temperatures and, in addition, are subjected to varying loads and infrequent starts. The lubricant in an engine serves to cool rubbing surfaces and provides a hydrodynamic film to prevent metallic contacts. In addition, it carries away products of combustion from combustion chambers and removes metallic and abrasive products. In order to insure satisfactory performance in a variety of engine designs under widely diierent operating conditions, natural petroleum products are specially compounded with oxidation and corrosion inhibitors, antifoaming agents, detergents, dispersants and other additives to produce the desired lubricating oil properties. Oils are qualified by running laboratory tests, both in and out of operating engines. A good brief discussion of these tests can be found in SAE Information Report J304a [16]. Based on tests such as these, oils have been classified by the American Petroleum Institute as to their suitability for use in engines under operating conditions of differing severity. For gasoline engine use, oils are classified in order of their ability to cope with increasingly severe operations as ML, MM, and MS and for diesels as DG, DM, and DS. I n addition, there are numerous military specifications and commonly used descriptors which cover the same basic oil properties [23]. In general, the severity of engine operating conditions and the design of the particular engine will determine the proper lubricating oil to be used. Sustained operation a t high load is not the only condition which may be called "severe." In fact, other conditions such as high sulphur or carbon content of the fuel, widely fluctuating loads or ambient conditions, frequent starts and stops, or atmospheric contamination may impose more severe oil requirements than high loads alone. Approximate military specification equivalents to commercial lubricating oils DG, DM, and DS oils are MILL-21044, MIL-G2104B, and MIL-L-45199 respectively. MIL-L-9000 is a Navy specification oil with increased resistance to the deteriorating effects of water contamination. It is below MIL-L-45199 in detergency level. The best judge of the proper oil to be used in an engine is the engine itself. Where past experience with a particular engine or with special operating conditions is unavailable, the judgment of the engine manufacturer and oil supplier must be relied upon. 4.4 Lubricating Oil System Design. The components of the lubricating oil system are usually furnished by the engine manufacturer and, in many cases, are completely assembled to the engine for installation in the ship as a unit. An oil sump is usually located under the engine and a positive-displacement pump takes suction from the sump and &scharges t h e oil to the engine through a flter, cooler, and strainer, in that order. The

I

MARINE ENGINEERING

MEDIUM AND HIGH-SPEED DIESEL ENGINES

1

I

EXTERNPLL RELIEF

METAL EDGE STRAINER

Dlesel engine lubricating oil consumption will vary depending on engine speed, size, and design details. Typical values of oil consumption are:

DUPLEX PRESSURE

II

COOLER BY-PASS

Medium-speed engines. . .3000-6000 bhp-hr/gal High-speed engines. . . . . .2000-3000 bhp-hr/gal

I - - - - - LUBE OIL - - - - - --,COOLER

\PRESSURE PUMP

,

ENGINE JACKET WATER (A) LUBRICATING OIL SYSTEM W I T H A FULL-FLOW F I L T E R AND WET SUMP

SCAVENOING

I

PRESSURE PUMP TO FILTER, ETC.

( 0 ) LUBRICATING OIL SYSTEM WlTH A DRY SUMP

M

I ------,

L---

JACKET WATER

FINE FULL FLOW

-

STRAINER LUBE OIL COOLER

---J

--

(C) LUBRICATING OIL SYSTEM WlTH A BY-PASS FILTER

-(D)

TO COOLER AS IN(A) OR(C)

ALTERNATIVE SYSTEM WlTH A SHUNT FILTER Fig. 15

Diagram of various lubricating oil systems

pump is equipped with a pressure relief, or in some instances, a pressure regulating valve. This system is shown diagrarnaticdly in Fig. 15(a). The sump tank should be sdiiciently large so that the oil does not splash up to the level of the crankshaft seals and so that the connecting rods will not dip into the oil under pitching an4 rolling conditions. I n addition, the sump should contain a quantity of oil in gallons equal

to about twice the rated capacity of the presswe oil pump in gallons per minute. If space is not available, the sump capacity may be less but not below a one-half minute pump supply. However, under these circumstances, oil change periods will be shortened appreciably. A much preferred solution to the problem of lack of space under the engine is to use a dry sump installation as shown in Fig. 15(b).

Oil change periods will vary with the severity of engine operation, quality of the lubricating oil, and size of the sump tank. With a dry sump, it is necessary to provide an additional pump to move the oil from the oil pan to the sump tank. This scavenging pump should have a capacity a t least 25 percent greater than the pressure pump to insure that the dry sump will, in fact, be dry. The oil flow requirements of engines will vary considerably, depending on such things as the use of oilcooling for pistons, whether the engine is naturally aspirated, supercharged, or after-cooled, and whether it is a two- or four-stroke cycle. The oil pressure pump capacity can be estimated a t about 0.2 gpm per horsepower for preliminary sizing of the system, though it might be half as much for some engines. Pump discharge pressures up to 100 psi can be expected in some engines. Since marine engines may run a t low speed for pr+ longed periods, engive-driven lubricating oil pumps should have adequate capacity to provide pressure under these conditions. Normal practice is to provide fullspeed pressure a t one-third speed. Many engines designed for constant-speed generator drive are found to be inadequate in this regard. ABS rules [24] require that the lubricating oil piping be entirely separate from other piping systems. For other than automotive-type engines, it is good practice to include a motor-driven lubricating oil pump in the system to be used to prime the engine before starting. The motor-driven pump is sometimes installed so that it can circulate oil from the sump tank through a heater and filter and then back to the sump in order to purify the oil while the engine is not runtiing. If this is done, care must be taken to insure that the normal oil supply to the engine can never be blocked off by negligence in realigning the valvivg prior to an engine start. It is possible to overprime opposed-piston engines, and the manufacturer's recommendations regarding means to prevent damage from this cause should be followed. Normally, the ABS rules require that an independently driven lubricating oil pump be furnished. However, for vessels in river or harbor service or vessels below 300 tons, this requirement is waived. I n those applications where the size and design of the engine is such that lubrication before starting is not necessary and an attached pump is normally used, an independently driven spare pump is not required if a complete duplicate of the attached pump is carried as a spare. Lubricating oil must be kept clean and free of abrasives. The best way to control abrasives is to prevent their entrance into the lubricating oil system. The designer should insure that filler caps are provided and located so

267

that foreign matter cannot get into the system n-hen it is being filled. Provision must be made to prevent dead pockets where deposits can accumulate and subsequelltly break loose in large quantities and cause damaging wear. Clean-out openings must be provided a t all locations where sludge is likely to accumulate. The diesel engine lubricating oil must be kept free of abrasive and corrosive q-mterials if it is to function properly. Additives are used to control corrosion, and filters are used to control abrasives. There are three commonly used filtering arrangements: (a) full flow, (b) by-pass, and (c) shunt. Full-flow filtration has become predominant in recent years, and, as its name implies, all of the oil supplied to the engine goes through the filter. This arrangement is shown in Fig. 15(a). Inasmuch as all of the oil going to the engine passes through the filter, it is necessary to prevent oil starvation of the engine in cases of filter plugging. An external by-pass line around the filter, together with a pressure relief valve, provides this protection. The duplex pressure gage shows the inlet and outlet pressures and gives advance warning of impending filter clogging. Normally, this takes place slowly so as to enable filter element changes to be scheduled during nonoperating periods. With the arrangement shown, the pressure relief valve setting can also be checked by means of the duplex gage. The lubricating oil cooler is installed after the filter because it is more effectiveto filter hot oil, as the pressure drop through the filter is less and filteeng is more complete. The simplex metal edge strainer is installed as close to the engine oil manifold inlet as possible to prevent the entrance of foreign matter into the engine. A by-pass filtering system is arranged as shown in Fig. 15(c). In this case, the oil discharged from the pressure pump is divided into two streams; one goes to the oil cooler and thence to the engine, and the other goes through a flow controlling orifice to the filter and thence to the sump. The quantity of oil by-passed through the filter to the sump must be in excess of engine lubricating requirements. The full pump discharge pressure is available for the pressure drop across the filter and orifice. By-pass filtration flow is approximately 5 percent of the pump A shunt capacity. filtering system is shown in Fig. 15(d). In it, the full flow to the engine is made up of oil which flows through the shunt filter and oil which flows in a by-pass around the filter, the quantity of by-passed oil being controlled by an ~rifice. There are three types of filter elements: those made of fine-mesh wire screen or metal edge (such as stacked disks); absorption types which are made of wool or cotton yarn, cellulose, or impregnated paper; and adsorbent types which, by adhesion, hold molecular layers of the contaminants to the filter element. The adsorbent elements contain fullers or diatomaceous earth, chemically treated papers, charcoal, or active clay. These filters are capable of removing additives from oil and should not be used with detergent lubricating oils except as part of an oil reclaiming system which is run separately

268

MARINE ENGINEERING

from the engine oil system. Additives should be restored to the oil after reclaiming and prior to reuse.' To provide an indication of the size of full-flow oil filters, the dimensions and flow rates of elements covered by specification MILF-20707 are given in Table 5. Table 5

Class 1

Characteristicsof Full-flow Oil Filters Max Dia (in.) 3

Maximum Instatled Length (in.) 4

A prox. Flow h t e (gpm) 2

Filter elements may be contained in individual containers, though it is more common to install several elements in one filter case. The elements may be stacked two or more high and arranged in any desired pattern in order to shape the case to suit available space; however, a cylindrical case is most common. It is essential that relief valves not be installed at the bottom of the filter case, where foreignmatter accumulates only to be washed into the engine whenever the relief valve lifts. Lubricating oil coalers are generally of the shell and tube type. For compact units, tubes may be fabricated in other than cylindrical form and include extended surfaces to increase heat-transfer rates. It is recommended that the pressure drop on the oil side not exceed 10 psi a t operating temperatures and that on the waterside be limited to 5 psi. The lubricating oil should be cooled with fresh water, even though it results in a larger cooler. The benefits in terms of faster oil warmup, reduced waterside fpuling, and better temperature control will more than offset this size increase. 4.5 Cooling Systems. As is true in all heat engines, the diesel engine must reject heat to the environment. Quantitatively, this heat is equal to the difference between the heat released by the injected fuel and the work output. The rejected heat is in the form of heat in the exhaust gas, heat transferred to the cooling system and lubricating oil, and the loss to the atmosphere due to radiation and convection from the engine exterior surface. It was previously stated that diesel engine efficiency is now being calculated and reported on the basis of the lower heating value (LHV) of the fuel in order to be consistent with presentations for other heat engines. However, much of the heat balance data in the literature, when reported on a percentage basis, will be found to be based on the higher heating value (HHV) of the fuel. For many years, the standard rule of thumb for estimating diesel heat losses has been, "One third of the heat in the fuel is converted to work, one third is lost in the exhaust gases and radiation, and one third to the cooling system. " The modern medium and high-speed, highoutput engines are more efficient than older engines and rather than one third of the input heat being converted to work, it can be expected to range'between 35 and 38 percent HHV (38 to 41 percent LHV), while about 28

percent is rejected to the cooling water and lubricating oil. While these percentage figures are of historical and general interest, figures in terms of Btu per horsepower per minute gre more useful in design work for estimating sizes of coolers, ventilation heat loads, and piping sizes. Average values for these heat losses are: To cylinder jackets. ....... .20-30 To oil coolers. . . . . . . . . . . . . . 5-20 To air coolers. ............. 5-10 To exhaust ................ 25-40

Btu/hp-min Btu/hp-min Btu/hp-min Btu/hp-min

These values will vary with engine design, load, speed, temperature of the coolant and oil, and degree of supercharging and aftercooling. For preliminary design purposes, the higher values may be used and about ten percent should be added when sizing coolers. After a specific engine is selected, exact values will be furnished by the engine manufacturer. In order to properly size the cooling system for an engine, the manufacturer must be provided with information relative t o the expected ambient conditions under which the engine will operate. For naval ships which must operate in widely varying locations, as an example, it is specified that coolers should be sized on the basis of an 85 F seawater temperatuye. If it is known that the ship being designed will operate in colder water, the cooler size can be reduced; or, on the other hand, if due to peculiar conditions ambient cooling water temperatures are exceptionally high, larger coolers will be needed. The discharge temperature of the seawater from the coolers should be kept well under 130 F to prevent scaling of the surfaces. Engine manufacturers design their equipment so that the water temperature rise of the fresh water across the engine will be between 10 and 20 deg F. This is done to minimize thermal stress and distortion in the engine. The capacity of freshwater pumps is usually in the range of 0.3 to 0.5 gpm/hp. The capacity of seawater pumps should be the same in order to simplify manufacturing and repair parts stocking, provide a margin to accommodate additional equipment such as aftercoolers on turbocharged engines, and prolong seawater cooler cleaning intervals by minimizing the seawater discharge temperature. A typical cooling water system for a medium-speed marine diesel is shown in Fig. 16. Automotive-type marine diesels usually are supplied with all piping, coolers, thermostatic valves, and expansion tanks assem-' bled to the engine. In this case, the only water connection the shipbuilder is required to make is from the sea to the seawater pump suction. The seawater pumps are likely to be subjected to reduced pressure a t the inlet, so to prevent loss of suction it is recommended that pump seals be of a type which will prevent air leaking into the pump under a suction head of 15 f t of water. The expansion tank should be located a t the highest point in the system and all pockets should be vented to the expansion tank. Water piping should be shed to match the pump suction and discharge flanges, or at least

TO WASTE HEAT RECOVERY SYSTEM

CI

G L t N G

THERMOSTATIC TEMPERATURE REGULATING VALVE WlTH BUILD-IN MANUAL CONTROL

cow. 1. D . OF RETURN TO BE APPROX. 3 x I.D. O f VENT LINE

PILLARY TUBE

CTUATING BULB ENT LINES CONNECTED TO IGHEST POINTS OF ENGINE

TO OVERBOARD OR TO WET TYPE MUFFLERS

THROTTLING VALVE TO BE APPROX. 10 PIPE DIAMETERS

JACKET COOLER TO L.O. SYSTEM VP

NOTES

I TO BE INSTALLED ON ENGINE GAGE BOARD. 2 DRAINS TO BE INSTALLED IN LOWEST POINT IN JACKET WATER AND SEA WATER SYSTEMS. USE GATE VALVES. 3 EXPANSION TANK SHALL BE LOCATED IN THE SAME COMPARTMENT WlTH ENGINE. 4 SEA WATER SUPPLY FOR GENERATOR AIR COOLERS WlTH THROTTLING VALVE TO BE PROVIDED ONLY WHEN REQUIRED. 5 SEA WATER PUMP SUCTION PIPING TO BE OF SUFFICIENT SIZE, AND ARRANGED TO LIMIT VACUUM AT PUMP SUCTION TO 6' HG AT RATED RPM. 6 JACKET WATER BY-PASS ACROSS BOTH COOLERS SHALL BE PROVIDED WHEN REQUIRED TO OBTAIN SPECIFIED OPERATING TEMPERATURES 7 TANK FOR INITIATING AND MAINTAINING JACKET WATER TREATMENT. CAPACITY TO BE . -11/2OALLONS FOR EACH 100 GALLONS IN ENGINE SYSTEM. TO USE.CLOSE V A L-V .. E ~ ~ ~ ............ TO TANK AND OPEN VALVES ' C m 8 ' D ' T 0 DRAIN TANK. CLOSEm~..FlLL TANK WlTH CORRECTAMOUNT OF SOLUTION. CLOSE -C:OPEN~A:CIRCULATION OF JACKET WATER WILL FEED SOLUTION INTO SYSTEM. TEST SAMPLE FOR CORRECT CONCENTRATION. ~

kg.16

Diagram of typical cooling water system

to provide smooth transitions if the piping must be smallei-. It may be possible or desirable to replace the seawaterto-freshwater heat exchanger with a hull cooler in cases where the seawater is contaminated or weed-infested. The hull cooler may consist of pipes with gxtended heattransfer surfaces 'hlounted outside'the hull, or simply tanka inside the hull wherein the heat is transferred directly to the sea through the hull plating. Kort nozzle shells have been used in the same manner. Thermostatic valves should be used to automatically regulate the outlet temperature of the jacket water. The outlet temperature should be kept in the range of 160 to '185 F to minimize the size of coolers and to prevent corrosive cylinder wear [25,26,27]. It is recognized that operating personnel prefer to operate cooler engines, as less time is required for cooling down if repairs are

necessary, surface temperatures are not uncomfortable to the touch, and machinery spaces are cooler. It is important, therefore, that the system be designed in such a t a y that the desired operating temperatures cannot be altered easily by the operating crew. Thermostatic controls should be such that adjustment out bf the proper range is impossible and orifice plates should be installed in piping systems once the proper balance is established. The jacket water of diesel engines must be treated to prevent corrosion and to minimize the effectsof cavitation on cylinder liners and jackets. A number of cooling water treatments, including alkaline chromates, soluble oil, sodium boron nitrate, and sodium nitrate-nitrite, are used. The engine manufacturer should recommend the coolant best suited for his engine. It may be desirable, however, for large fleet operators to standardize the

MEDIUM AND HIGH-!SPEED DIESEL ENGINES

coolant treatment used in their fleets, in which case the engine manufacturer should be informed of the preferred treatment. No water treatmeht will last indefinitely; the water must be tested regularly and chemicals occasionaily added. To insure that this is done, provision should be made for drawing of samples from convenient locations and to provide easy access to chemical addition points. A filling funnel located against the overhead in a hot engine room is almost certain to result in neglect of water treatment. For boat engines or other engines which may be exposed to freezingtemperatures, conventional inhibited ethylene glycol antifreeze solutions should be used. Where engines are installed high above the waterline, or where a source of raw water is either not available or unsuitable, air-cooled engines should be considered. It is important to insure that the cooling air is a t a sufficiently low temperature and that the air supply is not restricted by inadequate grill or duct openings to the *weather. The wind direction and velocity should be investigated to ensure that they will not oppose the cooling fan and impede airflow. Direct air-cooled engines are somewhat noisier than the liquid-cooled engines itlasmuch as they have no water jackets around the cylinders to attenuate vibration and noise. This fact should be considered when locating the engine in the ship. Direct air-cooled engines are delivered complete with cooling fan and the necessary cowling. The ship designer must insure that the air gets to the cooling fan and that the hot air from the engine is discharged from the compartment and is not allowed to recirculate back to the fan suction. About 50 cfm/hp of free air is required for air cooling. Air cooling can also be applied to liquid-cooled engines by the use of radiators to transfer the heat from the jacket water to the air. There is somewhat greater flexibility in installing an indirectly air-cooled engine than there is with one cooled directly with air. It is possible to place the radiator remotely from the engine to optimize installation arrangements. The radiators may be installed horizontally or vertically. Care must be taken to insurge that the engine-attached water pump characteristics match the cooling system requirements and provide an adequate flow of water. As with the directly cooled engine, particular attention must be paid to avoiding restrictions in the airflow path and to prevent recirculation of the cooling air. Thermostatic control can be applied either to the waterside, in which case the thermostatic valve directs the water flow through or around the radiator core, or to the air side, in which case the thermostat may operate a valve to divert air around the core, vary the speed of the fan drive, or change the pitch of the fan blades. The last two are more efficient as fan power is minimized at light load or when the air is cold. 4.6 Waste-Heat Utilization System. It is possible to utilize the waste heat from an engine by schemes which range from the simplest of using radiated heat to keep the engine room warm to complex schemes for generating

steam and power from the steam. The two most common uses for waste heat are: (a) heating water which can be used to heat spaces, heat fuel, cargo, or to distill fresh water; and (b) generating steam for use in absorption refrigerntion plants, space heaters, distillers, heat exchangers and low-pressure steam turbines. Almost 100 percent of the heat rejected to the jacket water and lubricating oil and about 60 percent of the exhaust heat are economically recoverable 1281. The amount of heat recoverable depends on the system used and the extent of the recovery equipment employed. The quantity of heat available depends on the design of the particular engine and operating conditions. Average values for heat losses were previously listed; these values can vary considerably, even for the same engine design. For example, in the case of a Fairbanks Morse 38D 8 diesel [29], the heat rejection rate to oil and water has been found to vary: (a) From a minimum of 35 Btu/hp-min a t 720 rpm to 41 Btu/hp-min at 900 rpm. (b) From 36 Btu/hp-min with an oil outlet temperature of 170 and water outlet temperature of 165 to and water at 230 32 Btu/hpmin with the oil at 185 (c) From 36 Btu/hp-min a t full load to 159 Btu/hp-min at 25 percent load. (d) From 36 Btu/hp;min a t full load without turbocharging to 22 Btu/hp-min with turbocharging. The question of whether to use waste heat and how extensive a waste-heat recovery system to design is largely one of economics. A detailed study must be made to develop load-time cycle data so as to determine how much heat is available. At the same time, the demand for waste heat must be analyzed as well, to make sure that there is sufficient heat available to meet the demand a t the time it is needed and that the heat generated can be used. This section briefly covers the basic systems used to recover waste heat, giving the basic engine input data required by the designer to size the equipment to utilize the heat and to devise special arrangements to suit each ship design. The major heat recovery systems are: (a) Engine radiator to air. The air temperature leaving the radiator is between 100 and 150 P and can be used for preheating boiler combustion air or space heating. (b) Normal-temperature, hot-water systems. These use a normally closed system with a thermostat to control the water outlet temperature and a heat exchanger to transfer unused heat to the seawater coolant loop. Hot water to the waste-heat utilization loads would be taken from the system at point F in Fig. 16 and returned a t point E. Additional heat can be added to the water by an exhaust gas exchanger. I n this system, the maximum jacket water temperature is that which can be obtained without pressurizing the expansion tank and will range from 180 to 220 F. A variation of this system is to use a secondary circuit to transfer the waste heat to the utiliring equipment. A heat exchanger is used to transfer the jacket water heat to the secondary circuit, and the

1

f

1

1 1

I

exhaust gaa heat recovery unit, if used, is installed in the secondary circuit. (c) High-temperature; hot-water systems. This system uses jacket water engine outlet temperatures in the range from 220 to 250 P and functions essentially the same as the normal-temperature, hot-water system except that a higher pressure is required in the circulating systems, especially in the engine coolant circuit. I n this system, a pressure control must be provided in the engine coolant circuit which will assure a pressure a t all points in the system sufficiently high to prevent the formation of steam. The source of this pressure may be a static head imposed by an elevated expansion tank or controlled air pressure in the expansion tank. For 250 F water, a pressure of about 20 psig is required a t the engine. In this system, all circulating pumps must be suitable for the higher pressures and temperatures. Engine and piping system gaskets and seals must also be suitable for the imposed conditions. With this high-temperature cooling system, it will not be possible to cool the lubricating oil with jacket water. The heat from the oil cooler must be disposed of in a separate system if it is not possible to use it for preheat in some part of the wasteheat utilization circuit. It may mean that more heat can be abstracted from a normal-temperature system using the heat from the oil rather than from a hightemperature system which does not use this heat. Thermostatic controls must be provided to prevent exceeding the maximum permissible temperature and pressure controls to prevent boiling. Exhaust gas heat may- be recovered in the high-temperature system as well as in the normal-temperature system. (d) Hot-water and steam sgstem with a $ash boiler. This system is quite similar to the high-temperature, hot-water system with the expansion tank replaced by a flash boiler. The pressure in the boiler is lower than that in the hot-water system expansion tank so that the hot water can flash into steam. This type of system is usually designed to operate with a steam pressure of from 2 to 8 psig. The operating pressure is dependent upon the maximum design engine coolant temperature and is set so that the total pressure a t the engine outlet due to the steam pressure and the static head will prevent boiling in the engine jackets. As in the high-temperature, hot-water system operating a t 250 F outlet temperature, a pressure of 20 psig a t the engine is required. If the jacket water leaves the engine a t 250 F and 24 psig and the static head is reduced to 8 psig, an equilibrium condition will be established in the flash boiler with about 0.985 lb of water a t about 235 F being returned.to the engine and about 0.015 lb of saturated steam going to the waste-heat utilization system for each pound of water entering the flash boiler. The 0.015 lb of condensate returned from the waste-heat system is mixed with the water in the flash boiler prior to recirculating through the engine. Using a water pump capacity of 0.3 gpm/hp, the 0.015 lb of steam per pound of circulating water equates to about 2.25 lb of steam per hour per horsepower. I n this system, it is necessary

271

to provide a water level control in the flash boiler and to supply make-up from the condensate return system. Boiler pressure control must be provided to prevent the pressure falling to the point where boiling will occur in the engine jackets. All piping from the engine to the boiler must pitch upward. (e) Ebullient system. An ebullient system may appear attractive where steam is required a t pressures of 12 to 15 psig for use in absorption refrigeration or airconditioning systems or other applications [30, 311. I n the ebullient system, boiling occurs in the engine water jackets. The engine circulating water pump is removed and the flow is maintained due to the diqerence in density of the steam-water mixture a t the outlet and the solid water at the inlet to the engine. System performance is sensitive to restrictions in the cooling water system and to the slope of the cooling water line. Pitch and roll can disturb the flow of cooling water t o the engine. A temperature difference across the engine of about 2 to 3 deg F will be maintained. The steam-water mixture from the engine flows to a steam separator above the engines. The steam pressure must be regulated a t the separator to insure that the pressure does not become too low, causing excessive boiling in the engine jackets, or too high, resulting in an excessive outlet temperature from the engine water jackets. Exhaust gas boilers can be provided with the ebullient system either built into the steam separator or operating in parallel with it. With an engine outlet temperature of 250 F, steam is generated at the saturation pressure of about 15 psig rather than a t 8 psig as in the previous example using a flash boiler. An estimate of the steam production capability of the ebullient system is given in Table 6 €301. Based on the data given in reference [4], the steam production capability of exhaust gas boilers is approximately as given in Table 7. Table 6 Steam Production Capability of an Ebullient System

Type of Diesel Engine

Fuel Heat ~ ~ Btu/hphr

%cycle turbocharged non-

8200

4gde I naturally aspirated &cycle turbocharged

Lb Steam/bhp-hr at Water Jackets with RBted Load ~ Exhaust ~ Manifolds t , ~ ~ Air-cooled Waterc Recovery cooled Unit 1.65

1.95

1.10

8500

1.90

2.35

1.25

7300

1.10

1.35

1.20

N ~ EThe : above data are based on 0-psig steam and 100 F ambient. 85y0 magnesia System or equal. piping is considered to be insulated with 1 in.-

Table 7 Steam Production Capability of Exhaust Gas Boilers

nT" "'

Diese Engine %cycleengine 4cycle engine

Steam Production Caqabilities, Ib/hg-hr 5 P@ 10 pslg 15 ps~g 0.75 0.78

0.70 0.75

0.68 0.74

h

MARINE ENGINEERING

In all waste heat utilization schemes, provisions must to duct the engine air from the outside directly to each be made to cool the engine when the waste heat cannot be engine and provide a three-way valve to permit the utilized. Where steam is generated, it is necessary to engine to take air from the engine room or the weather. Each engine should be provided with its own exhaust provide condensers and feed pumps and to insure proper system. If space does not permit such an arrangement treatment of the make-up water. Fortunately, the water treatments required for boilers and engine water jackets and it is necessary to combine the exhaust ducts from are compatible [4]. This treatment would include a several engines, it is necessary that valves be provided water softener to give zero hardness and a pH value in the branch from each engine to prevent backflow into between 9 and 11. Exhaust gas boilers may be combined an idle engine. The size of intake and exhaust ducts may be estimated with mufaers and may also be provided with supplementary oil firing to insure a steam supply under all using a figure-d 3.5 cfm/hp for'the intake air and 8.5 engine load conditions. The engine exhaust tempera- cfrn/hp for exhaust gas. $hese values may be high for ture conditions must be acceptable to the boiler supplier. most naturally aspirated engines and some turbocharged 4.7 Intake and Exhaust Systems. The intake and engines; however, it is desirable to provide some margin. exhaust system consists of the piping, filters, and silencers Duct velocities of 100 fps for the inlet and 150 fps for the necessary to conduct the outside air to the engine and to exhaust are suggested for preliminary design purposes. lead the exhaust gas from the engine to the atmosphere. When a particular engine has been selected and its To perform effectively, the depression in inlet air actual air and exhaust requirements are known, duct pressure and the elevation of the pressure a t the exhaust sizes can be calculated to meet the allowable pressure outlet must be minimized. The correction factors of drops. If it is necessary to reduce the duct sizes, higher Table 1 indicate the effect on engine power output as the gas velocities may be used [32]. Contrary to what may be believed, a marine engine is pressure st the engine inlet is reduced and as the temperature a t the inlet is raised. Turbocharged engines not always 'provided with clean air, particularly in river are particularly sensitive to intake air pressure and and harbor operations and sometimes in offshore exhaust back pressure. It is recommended that the operations in the vicinity of a desert. In these cases, it exhaust back pressure a t the engine outlet be limited to is necessary to provide air filters or cleaners to remove about 12 in. of water for turbocharged engines and twice abrasive or oily particles from the air. There are that for other engines. The inlet pressure drop in the basically three types of air filters or cleaners: ducting should not exceed 6 in. of water. Excessive 1 Dry inertial. The air direction is changed in the pressure drops in the intake or exhaust systems or a high filter, causing the heavier foreign particles to be separated inlet temperature can cause a loss of power, poor fuel from the air stream. These filters may be of the cyclone economy, high temperatures of engine parts, jacket or impingement type. water overheating, and excessive engine deposits. 2 Dry paper. The air is passed through porous The inlet to the induction air system should be located treated paper which retains foreign matter. so that it is not possible to draw in engine exhaust gases, 3 Oil bath. The air stream is directed a t the surface hot air from ventilation system exhausts, spray from of lubricating oil in the sump of the cleaner. The air seawater, or flammable vapors from tank vents or other reverses direction at the oil surface, and picks up and sources. Flammable vapors are particularly dangerous carries "washing" oil to the filter media. Foreign matter as they can cause an engine to overspeed, and the normal is captured at the media and washed to the oil sump overspeed trip and fuel governor will be unable to shut where it can be drained. it down. Actual filters usually employ combinations of these Piping should be properly supported and provided with expansion joints to avoid strains on the engine manifolds three methods. I n addition, self-cleaning designs are or turbocharger flanges. The velocity through the cor- available. It is possible to obtain filters of reasonable rugated metallic hose type of expansion joints should be si3e with moderate pressure losses from a variety of specified to insure that the type furnished will be suitable. -sources [2, 33, 341. The installation should provide Condensate traps and drains a t the low points of the gages to measure the pressure drop across the filter to engine manifolds should be provided. Provisions should give warning of impending clogging. The filter must be be made for rain covers to prevent the entry of water into installed in a location where is can be removed easily for cleaning. This would appear to be obvious, but for idle engines. Engine air may be drawn from the engine room or some reason it is frequently overlooked when the details ducted directly from the atmosphere. It is simpler from of an installation are developed. Air intake silencers are necessary to prevent blower an installation standpoint to take the engine air directly from the engine room; however, this arrangement has the noise from creating uncomfortable conditions in the disadvantage that the space may be excessively cooled engine room or spaces adjacent to the air inlet ducts. in winter. I n addition, in summer or in hot climates, Positive-displacement blowers generate a low-frequency the air may be heated by other equipment in the space pulsation, whereas the noise from turboblowers is very and reach the engine inlet a t an even more elevated high in pitch and is more likely to be objectionable. temperature, resulting in a loss in power. It is preferable Engines are normally fu+shed complete with an air

273

MEDIUM AND HIGH-SPEED DIESEL ENGINES

I

I

$1

silencer for use when the engine draws air directly from the machinery space. If the air is to be supplied to the engine via a duct, this should be so specified in order to insure that the intake silencer will be suitable. Exhaust mufaers are provided to reduce the pulsations in exhaust line pressure due to the cyclic release of slugs of exhaust gas into the engine manifold as the exhaust valves in each cylinder open in turn. The m d e r also serves to reduce atmospheric noise a t the outlet of the exhaust system. M d e r s may be of the wet or drv type. wet mufflers are infrequently used except & small boats, as they are limited to horizontal installations where the exhaust is through the hull of the ship above the waterline. Seawater is injected into the m d e r and cools the exhaust gas as the water is vaporized. The steam exhaust gas mixture is discharged overboard. With a wet m d e r , care must be taken to insure that the exhaust does not blow across the deck or against the side of adjacent ships. They should be fabricated of AISI 316L stainless steel for a reasonable life expectancy. Dry-type mufflers may be installed horizontally or vertically in the engine room or in the exhaust stack. These mufflers should be provided with spark-arresting features to prevent hot carbon particles from impinging on topside surfaces. I n general, mufflers should be capable of reducing the overall noise of exhaust gases to a maximum of 92 db a t a radius of 10 f t from the end and 2 ft above the muffler tailpipe with the engine operating a t rated load and speed. The noise level permitted may be more or less than this, depending on the particular installation. Figure 17, which was talcen from reference [35], gives an indication of the weight and size of naval dry mufflers with spark arresters. The muffler inlet flange size is the same as the exhaust pipe size. The pressure drop through these mufflersshould not exceed 18 in. of water a t engine rated speed for nonturbocharged engines and 6 in. of water for turbochafked engines. The pressure undulations in the exhaust from a turbocharged diesel are considerably reduced in flowing through the turbine to the extent that a much smaller muffler is generally permitted. However, spark-arresting features are still required. The muffler is generally installed in the exhaust pipe about one third of the distance from the engine to the end of the pipe. This distance will vary with the type of engine, the type of muffler, and the piping arrangement. The precise location of the m d e r should not be fixed until the engine selection has been made. ' 4.8 Starting Systems. To start an engine, it is necessary to rotate it, such that its speed and, consequently, its compression temperature are sufficiently high to insure ignition of the fuel when it is injected into the engine cylinder. The starting system is the means of supplying the energy for rotating the engine. The starting system can be operated with air, electricity, or hydraulic fluid. Air can be applied directly into the engine cylinders or used to drive an air motor geared to the engine crankshaft through an overrunning clutch.

----

L

nh-+lacier

will trip on overcurrent, short circuit, or engine overspeed, and they can also be tripped manually by either the engine room or motor room operator. The overcurrent trip is set for currents higher than any peaks encountered in normal service, and protection against moderate current overload is not necessary since the engines cannot exceed their maximum rating long enough to seriously overload a generator. At low propeller speeds the engines are run a t their idling speed of 300 rpm and the generator voltage is controlled by varying the generator excitation. This is accomplished by means of a potentiometer rheostat connected in the field of the generator exciter which provides excitation for all generators on the same bus. This rheostat is connected mechanically to the governor control cam and both are driven by the speed control handwheel. A pilot motor driving the handwheel makes

it possible to operate the speed control from any desired remote location. Remote control of the propulsion motor speed consists basically of a system for remotely operating the master speed controller on the motor room control board. The master speed controller in the motor room is the device that actually controls propulsion speed a t all times. Regardless of where the remote control may be originating, the motor room operator can take over control a t any time merely by disconnecting the remote control signal. Rather comprehensive tests were made during the trials of the Glacierto show the performance of the propulsion system. Oscillographic records were made during a series of maneuvering operations, and the data are summarized and reviewed in reference [4]. This interesting art5cle reveals a great deal about the operation of this well-engineered, -high-powered electric-drive system.

Section 3 The Turbine Direct-Curre~~tDrive System 3.1 System Composition. A turbine d-c drive system usually comprises a single, high-speed, nonreversing steam turbine driving a propulsion generator through a reduction gear, a propulsion motor, a control

system, connecting cable, and various auxiliaries. The turbine-generator set may be operated a t constant speed, and it may be convenient in such a case to add a ship's service generator to the same set.

MARINE ENGINEERING 11 0 0 1000 900

800 700 600

d

W

500

W

%

400

a

0

5

300

W 2

5 0 0 TO 1 0 0 0 VOLTS

1000

Fig. 11

.

2000 3000 4000 6000 GENERATOR RATING. KILOWATTS

10,000

Maximum d-c generator speed venur rating

The turbine-generator set requires the use of a reduction gear to reduce the inherently high speed of the turbine to a speed that is acceptable to the d-c generator. I n some cases it is considered advantageous t o utilize twin- or double-armature generators so as to take advantage of the higher speed a t which the smaller double unit can be run. Figure 11 shows the relationship between kw rating and maximum speed a t which normal d-c generators can be operated with reasonable design, performance, and maintenance. 3.2 Differences from Diesel Electric System. From an electrical standpoint, the operation of a turbine d-c electric drive is basically the same as that of the diesel system. Speed and direction of rotation of the propeller motor are controlled by varying the magnitude and polarity of the propulsion generator voltage through

The system differs in that it is usual to employ only one turbine per screw since it is'more economical to build a single turbine of full rating than a number of smaller ones. It is also possible and often desirable to operate the turbine-generator set a t constant speed and use the same turbine to drive a ship's service generator as previously noted. On high-speed diesel-electric systems, it is highly desirable from an engine-operation standpoint to reduce the engine speed whenever less than maximum power is required. 3.3 Applications. The turbine d-c drive is limited to cases where the extreme flexibility of the variable voltage control system is desirable, and where conditions are such that a turbine power plant is more desirable than a diesel plant as a prime mover. Applications of this form of drive are very limited. The U. S. Corps of Engineers has a number of turbine d-c electric dredges in operation but no new ones have been constructed in recent years. A noteworthy example of the application of,turbine d-c electric drive to a special service vessel is the cable ship Long Lines [5]. This large oceangoing vessel is a twin-screw ship with a total shaft horsepower of 8500. Each of its twin systems consists of a constant-speed steam turbine and reduction gear driving a doublearmature propulsion generator, an a-c ship's service generator, and an auxiliary generator which supplies power to the bow cable drive. The propulsion generator is of the double-armature type so that it can be operated a t the desired speed of 900 rpm. Each double-unit generator supplies the power required by its associated single-armature, direct-drive propulsion motor which is capable of developing a maximum power of 4250 hp a t 135 rpm. The a-c auxiliary generator and the bow cable drive generator are each rated at about 500 kw. The twin-screw turboelectric propulsion system provides excellent maneu~erabilit~y,speed control, and remote operation desirable during cable laying and recovery. The propulsion motors can be operated from the main motor room, three locations on the bridge, a location near the bow sheaves, and from the aft steering station.

Section 4 The Turbine Alternating-Current Drive System 4.1 System Composition-Synchronous. All of the discussion in this section on a-c electric drives will be based on the use of synchronous machines. These systems will invariably be of the three-phase type and operate a t unity power factor. They will not, however, always be 60-cycle systems since prime mover considerations may influence the use of a higher frequency and the propulsion system is normally independent of other power systems. Frequently, however, 60 cycles

per second is a good selection, and this facilitates the use of industrially equivalent designs [6]. I n the case of steam turbines, this form of propulsion makes possible the use of a single nonreversing highspeed turbine operating at or near its most efficient speed. The system normally comprises, in addition to the steam plant, a single, variable-speed, nonreversible steam turbine driving a direct-connected generator, a propulsion motor, a source of excitation power, control

'

ELECTRIC PROP'ULSION DRIVES

equipment, interconnecting cable, and certain auxiliaries. If the prime mover is a gas turbine, it will normally be of the two-shaft type so that wide speed range operation, essential for ship propulsion, will be possible. The power turbine output speed on large-size gas turbines that might be fitted to moderate and large-size vessels is in the 3600 to 5400-rpm range, and is thus similar to the normal speeds encountered on steam turbines of similar rating. The output torque-speed characteristics are also similar, and thus the a-c electric drive that would be employed is basically similar for either steam or gas turbine prime mover systems. In either steam turbine or gas turbine applications, the generator is normally of the direct-connected, high-speed turbo-type. The generator is similar to the ususl central station unit and uses a distributed field winding placed in slots in a cylindrical steel rotor. It will normally have two poles although four-pole units are also feasible. The rotor carries ventilating fans at each end which circulate air in a closed system through the machine and water-cooled surface air coolers. This type of totally enclosed machine is now universally employed and has the advantages of keeping the windings clean, simplifying the installation, and making a quieteroperating machine. The generator is provided with temperature-detecting coils inserted in several locations of the stator winding so that observations can be made of running temperatures. Arrangements are also made, usually by electric space heaters, for heating the machine when idle to prevent condensation of moisture. Figure 12 illustrates typical physical characteristics of a-c propulsion generators. The directdrive synchronous motor is of the salient pole type and is characterized by its large number of poles (60 to 72 being common) in order to operate a t the low speed required by the propeller. The motor is therefore large in diameter and short in stacked length. It must be provided with a heavier than normal pole face winding, so that it can operate successfully as an induction motor under heavy torque loadings produced by the propeller under reversing conditions with headway on the ship. Typical physical characteristics of a-c synchronous propulsion motors are shown in Fig. 13. Motors usually have forced-air circulation in a selfcontained system with surface air coolers. Because the rotative speeds are so low, motor-driven blowers are used to provide the necessary air circulation. These blower units are mounted on or adjacent to the main motor. The motor is normally provided with pedestal-type bearings bolted directly to the ship's.foundation, and one of these units can be combined with the propeller thrust bearing if desired. To facilitate major maintenance or repair, the span between motor bearings is increased sufficiently so that the motor stator can be shifted axially to expose the stator windings and the rotor pbles without necessitating bearing movement or shaft disassembly. This is an important feature that contributes

349

significantly to the ease with which inspections can be made and maintenance work accomplished with the motor in place. As in the case with generators, stator temperature detectors, fire-extinguishing connections, and space heaters are provided. The pxcitation requirements for the a-c system are considerably more severe and difficult to meet than for the d-c system. When starting or reversing the motor, it must operate as a squirrel-cage induction motor until its speed is electrically close enough to that of the generator so that it may be synchronized. During this out-of-synchronism mode of operation, the motor power factor is very low and its current demands high. To maintain generator voltage and provide the current needed to develop proper motor torque, the generator must be over-excited on a short-time basis. These conditions are particularly severe when the motor is reversed from a full-speed ahead operating condition because the ship continues to move through the water a t considerable speed and the water flow to the propeller causes it to resist motor efforts to stop and reverse it. Typical propeller torque-rpm characteristics are shown for various ship speeds in Fig. 14. With the a-c drive, it is necessary to have a separate source of excitation power. It is not practical to use generator-driven exciters because of the wide speed range of the turbine-generator set (100 percent to 20 percent speed) and the fact that the exciter must provide approximately 2.5 times normal voltage during maneuvering, which is done a t the minimum generator speed. Excitation is commonly supplied from separate motorgenerator sets operating from the ship's service system. A pilot exciter of the rapid response or of the static regulator type is also used so that regulating and limiting control functions may be provided conveniently. A standby excitation set should be provided to safeguard the availability of this vital auxiliary. I n the case of twin-screw ships, a total of three excitation sets would be supplied, the third unit being arranged as the spare for either of the two propeller systems. Under steady running conditions, sufficient excitation must be maintained on the main generator so that the main units do not pull out of step because of torque variations of the propeller caused by ,turning or sea conditions. , This can be done by maintaining a degree of overexcitation, or regulators may be employed which act automatically to provide increased excitation when torque variations require it. I n the turbine-electric a-c system, speed control of the propulsion motor is obtained by frequency control, or, in other words, by varying the speed of the prime movergenerator set. The turbine is under the control of . a governing system which permits its speed to be varied over the range of from about 20 percent to 100 percent speed. All steady-state running is performed with the main motor in synchronism with the generator and the speeds of the motor and generator proportional to each other.

350

MARINE ENGINEERING

ELECTRIC PROPULSION DRIVES C

REMOVE COOLERS

I

+ =PLAN A VIEW R O ~ RR&OVAL

Fig. 12

A-c proplion generatorsdimemions and weights

- -

HP

WEIGHT IPOUNDSI DIMENSIONS (INCHES I ISQRPM I A I B I C I D I E I F I G I H I I I J I

WEIGHT

3 5 0 0 0 ~ 2 4 8 ~ 2 8 4 ~ 2 8 5 ~ 1 2 5 ( 1 3 5 ~ 1 2 5 ~ 1 7 2 ~ 1 31620)080 2 0~74~ 14511351195 115) 701 821

450000

I 0 3 RPM WOO0 )31b(3281338 ~ l 5 8 ~ 1 6 7 ~ 1 5 6 ~ 2 0 06) 0 1 21 8 7 4~1

525000

r weigh Fig. 13 A-c propulsion m o t a r 4 i m ~ s i o n and

Figure 15 shows a schematic wiring diagram for a typical single-screw turbine-electric a-c propulsion system. The direction of rotation of the propulsion motor is controlled by reversing contactors (S1 to 85). This reversing switch group can be remotely controlled switches, contactors, or breakers, depending upon the control arrangement and system size. Excitation power is supplied by a main or standby excitation motorgenerator set driven from the auxiliary (ship's service) power system. Other excitation systems can be used, such as auxiliary turbine or diesel-driven exciter or static exciters supplied from the ship's service power system. The propulsion motor field is controlled by an excitation contactor with contacts E l , E2, and E3. The field of the propulsion exciter is automatically controlled by a propulsion regulator. This regulator automatically adjusts the excitation power level in response to the direction control lever and throttle control lever movement. The control station, depending upon the control arrangement, will have either one or two control levers. With a two-lever station, one lever controls the direction switch (S1 to 85) and supplies the logic signals for automatic starting and synchronizing of the propulsion motor. The other lever provides speed control logic to the prime mover governor. Interlocking between the two levers is provided to (1) allow moving the reverser lever only when the throttle lever is in the maneuvering speed position, and (2) to allow movement of the throttle levers from the maneuvering position only when the reverser is in the run position and propulsion motor has started and is synchronized with the generator. With a two-lever control station, the following sequence would be automatic in response to the lever movements: A. REVERSER lever is in STOP THROTTLE lever is in MANEUVERING 1. The turbine-generator is running a t minimum speed (15 to 25 percent rated). 2. Excitation bus is a t zero volts with the M-G set running. 3. Motor field is shorted through its discharge resistor ( E l and E2 open, E3 closed). B. REVERSER lever is moved to AHEAD 1. Direction contacts S1, 52, and 53 close. 2. The propulsion regulator increases generator excitation to maintain rated volts per cycle or maximum level required by system design, usually between 6 and 7 times rated generator field power. 3. Propulsion motor starts and accelerates as an induction motor. 4. The field of the propulsion motor is automatically excited by the closing of the motor field contactor. ( E l and E2 close and E3 opens). The automatic synchronizing system should be designed to apply motor field at the proper slip and a t a phase angle to obtain a high pull-in torque. C. THROTTLE lever is moved to the desired ' propeller speed 1. The turbine speed governor resets to call for the speed indicated by the throttle lever position signal.

35 1

Fig. 14 Typical propeller rpm-torque curves for various h i p speeds

2. Fuel-power increases until the system speed. reaches the turbine governor set point. The following sequence would be followed in reversing from full AHEAD to full ASTERN: A. THROTTLE lever is moved to the minimum speed position 1. The turbine governor is reset to call for minimum speed. B. REVERSER lever is moved from AHEAD to ASTERN 1. Excitation voltage goes to zero. 2. Motor field contactor shorts motor field ( E l and E2 open, and E3 closes). 3. Reversing switch contacts S1, 52, and 53 open and+contacts52, 54, and S5 close. 4. The propulsion regulator increases excitation to the generator to maintain rated volts per cycle or to the maximum excitation power limit. 5. The propulsion motor is now operating as an induction motor. It will reverse and accelerate to a speed close to the synchronous speed of the turbinegenerator. 6. The field of the propulsion motor is automatically excited and it pulls into step as a synchronous motor. C. THROTTLE lever is moved to the desired propeller speed

MARINE ENGINEERING

EXCITATION BUS

SHIP SERVICE POWER

Fig.

IS

Typical schematic circuit diagram-turbine electric a-c drive

1. The turbine speed governor resets to call for the speed indicated by the throttle lever position signal. 2. Fuel-power increases until the system speed reaches the turbine governor set point. Control systems with higher degrees of automation, programmed sequencing, and remote actuation can be employed, and the trend is to move in this direction. Protection against faults or short circuits in the main circuit is provided by phase balance relays or a differential relay, or both. The functioning of these relays causes excitation to be removed from the propulsion motor and generator. Ground protection is provided by a ground current relay circuit in the propulsion motor neutral. Functioning of this relay also removes excitation from the propulsion motor and generator. Systems which utilize multiple prime movers must incorporate means to obtain proper kw load-sharing between prime movers, kvar sharing between generators, and provisions for adding or removing a generator set from propulsion duty. Real load-sharing between propulsion engines is usually accomplished by speed droop. Since good speed regulation (low droop with load) is not required or even desirable, this droop can be set much higher than normally considered for ship's service generator sets. A speed droop setting of 10 percent should provide good load-sharing and stable operation. Reactive load-sharing between generators can be accomplished by connecting the generator fields in parallel from a single exciter, or by reactive droop if individual exciters are used. Various methods can be used for adding or removing generator sets, depending upon the type of prime mover and generator. The most straightforward method is to provide adequate switchgear and automatic paralleling between propulsion generator sets. Where multiple prime movers are used, it is desirable

to provide for operation on less than the full number of propulsion sets. For instance, if three turbine-generators are used to provide power to one propulsion motor, it is desirable to incorporate coritrol features that enable operation on 1, 2, or 3 generator sets. Operation on a reduced number of generators (reduced power capability) increases the propeller reversal time and electrical machine load because the propeller torque during a reversal from high ahead ship speeds exceeds the plugging motor torque available with reduced power input. The maneuvering (reversal time) performance is improved and overloading required of the generators and motor is reduced by the addition of a dynamic braking resistor. This resistor is connected to absorb the propulsion p u m p back power due to the forward motion of the ship and the resultant water action on the propeller. When the ship has lost sufficient headway, the dynamic brake is disconnected and the propulsion generators in use are connected in reverse phase rotation to the propulsion motor. The power from the reduced number of propulsion generators should now be sufficient to accelerate the propulsion motor to near synchronous speed in the astern direction. The general requirements for connecting cables as given in Section 2 for d-c propulsion apply. Cables should preferably be of the single-conductor type and should have nonmagnetic armor. The armor should be grounded at approximately the mid-point of the cable run. Single-conductor, a-c cables should not be located closer than 3 in. from parallel magnetic material and, where cables pierce a bulkhead, all conductors of the same set should pass through a common nonferrous plate to prevent heating of the magnetic bulkhead. Single conductor cables should be supported on insulators, and where cables are arranged in groups they should be transposed when lengths exceed about 100 ft.

ELECTRIC PROPIJLSlON DRIVES

Air coolers are supplied with the propulsion motors and generators and are normally built in as a part of the enclosed ventilating air system. The necessary motordriven vent fans for the main motor are also considered a part of the motor and are normally mounted as a part of the main motor. Lubrication of the generator bearings is provided from the turbine system. Propulsion motor ,bearings are normally force-lubricated from a separate system consisting of a motor-driven pump, oil coolers and strainers, and a sump tank. 4.2 System Design Features. For electrical losses in the transmission between the turbine and the propeller, an average figure of 6 percent can be taken. This figure does not include the excitation power of the generator and motor fields. This loss will amount to about 1 percent of the kilowatt rating of the generator and about 1.5 percent of the kilowatt rating of the motor. This loss is included separately since it is supplied from a source separate from the main turbine. The propulsion system, being an independent system, enjoys considerable freedom of choice as to voltage and frequency. As noted previously, the choice of voltage level is based on motor and generator machine design considerations, and on the availability of needed switchgear or control apparatus. As a guide, a 10,000-hp .system might well use about 2400 volts and a 50,000-hp system would find 6600 to 7500 volts advantageous. The minimum frequency of a turbine-electric, a-c system is largely determined by the speed of the turbinegenerator unit and the use of a two-pole generator design. Since the turbine will invariably have a rotation speed of 3000 rpm or more, this results in a minimum frequency of 50 cps. The corresponding motor, if its desired operating speed is 100 rpm, would have 60 poles. Higher frequencies would require more poles on the motor and tend toward a less satisfactory and less economical design. Frequencies, then, tend to be in the 50-90 cps range, with the lower frequencies favoring the slow-speed motor and the higher frequencies being accepted only as necessary from the turbine standpoint. The inherent torque characteristics of an a-c ship propulsion motor require coordination of the system design. The a-c motor and generator combination will not carry overload torques under steady running conditions a t full power unless the machines are designed with considerable torque margin or unless provision is made in the control and excitation system to automatically increase the excitation of the machines when the overload torque is imposed. Unless these provisions are made, the machines will pull out of step and require resynchronizing. The torque requirements on the motor during a fullpower, full-speed reversal are the most exacting and usually largely fix the design of both the motor and generator. Since synchronous motors are almost universally used in a-c propulsion plants, only this type of motor will be considered in the discussion of this problem. The rotor is provided with a substantial

353

induction winding which must be carefully designed so that it will not detract appreciably from the purely synchronous motor characteristics for steady running, and so that it will be able to develo~sufficient induction motor torque for stopping the propeller and bringing it up to speed in theastern direction while the motion of the ship through the water is still attempting to drive the ~ r o ~ e l l eand r the motor in the ahead direction. The ind;ction winding must be carefully proportioned so as to have sufficient thermal capacity to handle the heavy currents induced during the maneuvering cycle. In order to studv further the characteristics necessarv to accomplish a Ifull-power, full-speed reversal, it is desirable to review briefly the sequence of operations during the maneuver. On signal for full-speed astern, the operator first moves the turbine speed control lever to the IDLING position, which sets the governor for about 20 percent speed. He then moves the field lever to the OFF position, removing excitation from both the generator and motor. He next moves the direction lever from the FULL AHEAD position to the START position astern. While these operations have been going on, the ship has been slowing down because power was removed from the propeller. However, it is still moving ahead through the water at a considerable speed and thus driving the propeller and motor in the ahead direction. In the last o~erationjust mentioned. two of the three-phase connections between the generator and the motor are reversed; and approximately double excitation is applied to the generator field. Power therefore is applied immediately to the stator of the motor in the reverse direction, causing large currents to circulate in the damper winding of the motor and therefore developing a heavy torque which acts to stop the propeller and then to reverse it and bring it up astern, close to synchronism with the main generator. The operator then moves the control lever to the RUN position, which applies field to the motor, pulling it into synchronism with the generator as a synchronous motor and a t the same time reducing the generator excitation to normal. While the motor is operating as an induction machine, it has a rather low power factor and places a current demand on the generator of from three to five times normal. If no provision were made for overexciting the generator field, its terminal voltage would collapse, and the motor would fail to deliver the necessary torque. This is the reason that the excitation system must be closely coordinated with the machine design. The motor design must be carefully proportioned so that it will develop sufficient synchronizing torque to pull into step with the generator although the ship still is going ahead and causing ahead torque to be developed by the propeller. As each type of ship and propeller will have different maneuvering characteristics, the first step in determining the proper design of the propulsion motor and generator is to calculate the expected maneuvering chmacteristics. Figure 16 shows the full-power reversal speed-torque curves for a typical single-screw, turbine-electric-pro-

354

ELECTRIC PROPULSION DRIVES

MARINE ENGINEERING Table 1

virtually all cases the main motor will be located as far aft as possible. This results in important reductions in length of line shaft, shaft alley, and number of bearings. Motor foundations should provide for axial space in which to shift the motor stator to uncover the windings of stator and rotor without disturbing the motor rotor and its bearings. As in the case of any propulsion plant, a careful analysis of the complete plant must be made to ensure that there will be no dangerous critical speeds within the operating range. Although the electrical machines provide damping action, it is possible to obtain vibration frequencies due to a combination of the propeller and the synchronous motor which will resonate with the natural frequency of the system. It is also possible to reflect such pulsations back into the turbine-generator set. It is therefore necessary that a careful analysis of the entire system be made. 4.3 Physical Characteristics. Because of the great variation in requirements and arrangements, precise data on equipment weight and size cannot be presented; however, general outline dimensions are given in Fig. 12 for a-c propulsion generators of the turbine type, and Fig. 13 gives similar data for a-c propulsion motors of the direct-drive type. These figures also give overall weights for these machines. The speed chosen for these machines is arbitrary but nevertheless typical and illustrative of units that would be used. Control or excitation equipments are small and light in weight when compared to the main motor and generators and lend themselves to mechanical packaging that facilitates convenient installation.

Allowable Temperature Rises for A-C Machines, Deg C

-

INSULATION Ambient .................... Armature windings by imbedded detectors.. . . . . . . . Salient pole fields b resistance. ~urbine-typefieldsty resistance

PROPELLER SPEED, O/o RATED

Fig. 16 Typical propeller and motor torque characbrirtiu

I

I

pelled oceangoing vessel. Curve A shows the calculated maneuvering torque capacity of the propulsion plant. Curve B shows the propeller torque requirements during maneuvering, starting with the ship going ahead at full speed. As the ship slows down, the propeller torque is reduced so that there is a family of curves similar to curve B to cover various ship speeds (see Fig. 14). Curve C is a portion of one of these curves and represents the propeller torque-speed conditions existing when the motor speed has come within the range from which, at point Dl it can be synchronized with the generator. The motor can be synchronized with its generator when the slip has been reduced to about 2 percent or less, and this ability is enhanced by timing the application of motor field so that the phase of its angular slip position is most favorable for synchronization. During induction motor operation, the motor field will .be short-circuited on itself or through a resistor chosen to aid the synchronizing process and to keep induced voltages.within acceptable values. The size, weight, and cost of a-c electric propulsion equipment are increased if inherent torque margins are increased. The torque margin of a propulsion system is defined as the increase in torque, above rated torque, to which the system may be subjected without having the motor pull out-of-step with the generator. A torque margin of 10 percent is sufficient when combined with an automatic control means for raising excitation momentarily whenever torque increases occur, such ss can be

Class B 50

Class F 50

70 70 80

90 90 100

Class H 50

110 110 120

produced when maneuvering. I n the case of twin-screw vessels, a sudden hard-over rudder movement can appreciably increase the loading of the inboard propeller and, in such in~tallations,load limit as well as excitation increase may be desirable. The comments made in Section 2.2 regarding the availability of improved insulations and the demise of Class A insulation apply to a-c machines as well. Class B insulation systems are the usual standard but the use of Class F is increasing. Allowable temperature rises are given in Table 1 and more complete details will be found in reference [I]. Again it should be noted that the ambient of totally enclosed machines fitted with air coolers is that of the air delivered by the coolers and, if designed for less than 50 C air delivery, corresponding extra temperature-rise allowances can be made. Propulsion motors and generators are rated for the maximum full-power, continuous shaft horsepower requirements of the vessel. Electrical equipment so designed, with temperature ratings as indicated in the preceding paragraph, obviously can carry some overload without distress. Such overload will result in higher operating temperatures and some reduction in overall insulation life of the machine. The amount of such life reduction depends on the severity and frequency of the overloads and on the length of time of such service. Experience indicates that it is advisable to provide both the generator and motor with closed ventilating systems and water-cooled air coolers. The preserving of clean windings in such systems, and the elimination of extensive air duct systems, makes the overall cost of the closed system favorable. I n cases where added safeguards against air-cooler water leakage seem prudent, double-tube type air coolers are used. For installations where the propulsion motor is located in a room by itself, it is not necessary to provide a completely closed ventilating system on the motor itself. In this case it may be advantageous to provide for the ventilating fans to exhaust the warm air from the motor through the air coolers into the motor room. The cool air is then drawn back into the motor through openings at both ends of the motor. The machinery layout selected will give consideration to many factors and these will vary for different types of ships. The mechanical independence of the main motor and the turbine-generator set, however, is useful and in

'

355

TZSE-A1 Tankers Turbine Generator. . . . . . . . . . . .5400 kw, 3715 rpm 2370 volts, 3 phase 62 cycles, 1.0 pf Main M ~. . . . ~ . . . . .~. . . . ~. . .6600 . hp, 93 rpm 2370 volts. 3 ~ h a s e 62 cycles, l.oSpf T2-SE-A2 Tankers Turbine Generator. . . . . . . . . . . ,7650 kw, 3715 rpm 3610 volts, 3 phase 62 cycles, 1.0 pf Main Motor. . . . . . . . . . . . . . . . . .10,000 hp, 106 rpm 3610 volts, 3 phase 62 cycles,. 1.0 pf ,

This type of drive is discussed further in [8]. b. Passenger Vessel :Canberra The Canberra is a large passenger vessel built for service between England, Australia, and the West Coast of North America [9]. The choice of steam-turbine, a-c electric drive was based on a number of factors, and was no doubt influenced by very favorable operating passenger experiencevessels. of its owners with three other turboelectric

The Canberra has a twin-screw propulsion plant with a combined rated shaft horsepower of 85,000. Each screw is driven by a 42,500-hp double-unit motor at 147 rpm full speed. The motors are of the synchronous type and operate at unity power factor and a maximum voltage of 6000 volts. Each of the double-unit motors is supplied with power by a single generator directly coupled to a high-performance steam turbine. The generators are somewhat oversized in order to permit a single unit to 4.4 Applications supply the desired amount of power to both propulsion a. T2 Tankers. Over 500 turbine a-c electric drive motors for certain legs of the vessel's itinerary. It is tankers were built during the 19401s,mainly in response interesting to note that the tested efficiency of the main to World War I1 needs. Many were converted to other motors is over 98 percent throughout the power range service such as ore-carriers and self-unloading colliers [7]. of 50 to 100 percent and is 98.4 percent at full power. An unusual aspect of this application is the emphasis Many of these vessels were "jumboized" by adding placed upon quietness of operation. The Canberra is longer midbodies and in some cases only the stern sections were retained. Invariably the propulsion plant probably the quietest and smoothest-running ship of its was retained and, in many cases, overhauled and type in service and her electric drive, while contributing uprated. There were numerous instances' when the to her quietness directly, also enables the two propellers propulsion turbine-generator sets were used to feed to be run in synchronism and in the phase position that power ashore in times of disaster or other critical need. produces the greatest neutralization of vibratory forces; Popularity and longevity are not the main reasons for the 42-pole motors permit 21 different synchronized citing this application but rather the fact that the basic operation relationships shaft-to-shaft. I n addition to plant is typical of a steam-turbine, a-c electric drive for a the actual vibration reduction, the elimination of the single-screw ship. The plant consists of two boilers usual "twin-screw beat" is a particularly important which supply steam to a single propulsion steam turbine psychological improvement on a passenger vessel. The direct-connected to a two-pole generator, one 80-pole twin screws of the Canberra are normally run in synchrosynchronous motor direct-connected to the propeller nism with each other except when in confined waters shaft, and an excitation and control subsystem. The where maneuvering is expected and independent propeller control is important to the handling of the vessel. rating of the main units is as follows:

.

120

Section 5 The Diesel Alternating-C~rrrentDrive System 5.1

Differences from Turbine a-c Electric Drive.

Few diesel a-c electric drives have been built but they are of particular interest from an electric-drive . standpoint. The diesel a-c electric-drive system consists of a multiple number of diesel engines, each driving a directconnected, salient pole, a-c generator, a single slowspeed, direct-drive propulsion motor, an excitation subsystem, and a control system. The main motor will most likely be of the synchronous type, and subsequent portions of this section will be based on that type of motor. The fundamental principles of speed control and reversal of the propulsion motor are the same as for a turbine-electric drive. However, the torque characteristics of the diesel engine a t reduced speed are less favorable than those of a turbine ank there is risk of stalling the engines during critical maneuvering if proper control safeguards are not provided. I n addition, a number of generating units must be operated in parallel over a range of from 30 to 100-percent speed. These factors combine to require a carefully designed control system to ensure that the operating conditions can be met within the capabilities of the equipment. 5.2 Description of a Typical System. As a means of illustrating a diesel a-c electric drive, consider the Navy Submarine Tender Hunley. The Hunley is a diesel a-c electric drive of 15,000 shp [lo, 111. This application is typical of those special-service vessels in which use can be made of the prime mover generating sets for auxiliary

service when they are not required for primary propulsion. The propulsion plant consists of six 850-rpm diesel-generator sets rated a t 2655 bhp each with a 2000-kw generator output a t 3300 volts and 70.8 cycles. These six generators operate in parallel to supply power to the 15,000-hp direct-drive propulsion motor. The ship may be propelled by any combination of 3, 4, 5, or 6 generators, and up to 3 generator sets can be used to augment the ship's service power supply when tending power demands are heavy. The basic propulsion system is shown in Fig. 17. A single exciter is arranged to supply field power to all of the propulsion generators and the propulsion motor. The main exciter is under the control of a pilot exciter of the amplidyne type, and it in turn is controlled by the propulsion regulator. This regulator automatically maintains proper excitation for all operating modes. Any of the generators can be connected to augment the ship's service system for special duty such as cargo handling, pumping, or overside power supply. As shown, excitation for this mode of operation is from static exciters. Typically, each of the gix diesel-generator sets consists of a high-speed engine direcbconnected to a salient pole generator. The diesel-generator sets must be varied in speed in order to obtain propeller speed control. This can be done from 100-percent speed down to about 30percent speed under steady-state operating conditions. At this low speed, load transients must be avoided to prevent danger of engine stalling, and as a result maneu-

-

PROPULSION. GENERATOR EXCITATION BUS

Fig. 17 Schematic diagrum-dienl a-c electric drive 0

357

ELECTRIC PROPULSION DRIVES

MARINE ENGINEERING

I/

440 VA? POWER

1

I

I

I

I 1

AHEAD ASTERN PROPELLER RPM,XRATED

Fig. 18

Propulsion motor torque characteristics with power supply set for 40-percent astern rpm

0 0

vering operations are carried out a t higher engine speeds, i.e., 4 0 4 5 percent, so as to have increased torque capability and assurance of nonstalling performance. The propulsion motor is directly connected to the propeller shaft and is rated a t 15,000shp and unitypower factor. A voltage of from 3000 to 4000 volts is indicated for this machine. The motor is equipped with as heavy an amortisseur winding as is practical for a machine of this type without requiring an unbalanced design or reflecting adversely on normal synchronous motor operation. The motor is totally enclosed with surface air coolers and motor-driven blowers. A motor of this type is shown in Fig. 13. The motor torque characteristics under maneuvering conditions are of paramount importance. Figure 18 shows typical motor speed-torque characteristics at 40percent rated frequency (the assumed maneuvering speed selected for the diesel-generator sets). Curve A is the motor performance under full rated volts-per-cycle supply conditions, as would be expected on most industrial applications operating from an unwavering power supply. On a ship application, the motor must be started from a generator of equal capacity and the motor torque must be determined on this basis. The generator characteristics, motor characteristics, and excitation system must be selected so that optimum system performance is obtained. Curve B is the system motor torque exclusive of any prime mover limitation. This curve is based on the volt-ampere output from the generators with their field forced to about 2.5 times normal excitation. Producing this motor torque, however, requires the generator to demand more torque from the engine than is available over a part of the speed range. Curve C, a modification of curve B, shows the motor torque available within the limits of engine capacity. The torque demand of the engine has been reduced so as to just equal, but not exceed, the available torque. This is done by regulating the excitation on the

20

Fig. 19

40 60 80 SHIP SPEED, % RATED

100

Maximum operating speeds

generators in response to a speed signal from the enginegenerator set. Curves D and E are similar motor torque curves but with only half of the six dieselgenerator sets supplying power to the motor. The free-route propeller power-speed curves are shown in Fig. 19 as the usual cubic curve. To this has been added the power capabilities represented by 3, 4, 5, and 6 diesel-generator sets. The intersections of these curves represent maximum operating speeds for the various combinations. The propeller characteristics of importance during maneuvering operations are the dynamic relationships of propeller torque and speed for a series of vessel speeds. Such characteristics are shown in Fig. 14. The power requirements shown by Fig. 19 represent the steadystate duty as seen by the motor, and the curves of Fig. 14, when augmented by an appropriate acceleration component, represent the transient duty as seen by the motor. These then are the output requirements of the system, and all control and system designs are directed to meet these needs while staying within the basic capabilities of the primemovers, thegenerators, and the propulsion motor. The propulsion control is divided into three sections. One controls the two excitation motor-generator sets and provides for selection of the in-service set. Another section provides for generator control and connection to either the propulsion bus or to the special ship's service duty bus; and a third section provides for control of motor speed and direction of rotation. The propulsion motor speed and direction of rotation are controlled by three levers: REVERSING, FIELD, and SPEED. These, together with the major interlocking, are shown in the functional diagram of Fig. 20. The arrangement of mechanical interlocks will prevent damaging operation. I n general, the REVERSING

I

I

358

MARINE ENGINEERING ENGINE SPEED TRANSMIlTER MECHANICAL INTERLOCKING

I

I

AHEAD OFF DYNAMIC BRAKE ASTERN

It I

(

I

R VERSING 'LEVER

FULL

RUN NO. 2

I

I

FIELD LEVER Fig. 20

I

I

CEE~ (

Motor control leven

% OF RATED TORQUE

Fig. 21

Propulsion system characbristict-rtarting and mnning

lever cannot be moved unless the FIELD lever is in the OFF position and the SPEED lever is in the MANEUVERING position. The FIELD lever cannot be moved unless the REVERSING lever is in the AHEAD or ASTERN and the SPEED lever is in the MANEUVERING position. The SPEED lever cannot be moved unless the FIELD lever is in the RUN position. Each generator is controlled by a POWER SELECTOR lever having four positions: SPECIAL SERVICE, OFF, PROP 1, and PROP 2. This lever permits any generator to be added or removed from propulsion duty a t any time, and to be used for special service power supply when not used for propulsion. In order to start the plant, the generators which are to be used for propulsion (at least 50 percent of full capacity) are first switched from local control in the engine room to remote control a t the main propulsion control board. This connects the engine governors to a master trans-

mitter, and all engines being readied for propulsion service will run a t approximately the same speed. The generators are connected to the propulsion bus by moving their individual POWER SELECTOR levers to the PROP 1position. When the first POWER SELECTOR lever is moved to the PROP 2 position, sufficient excitation is applied to energize the bus and synchronize all of the connected generators. Each will in turn then be moved to the PROP 2 ~ositionand excited a like amount. The propulsion b;s is now energized by all of the in-service generators but at a reduced frequency and very low voltage. Moving the motor REVERSER lever to the AHEAD position connects the motor to the propulsion bus in the ahead phase rotation ready for induction motor operation with its field shorted through a discharge resistor. The motor may not start in this position because of the low generator excitation but when the FIELD lever is moved to position No. 1, approximately 250-percent generator excitation is applied and the motor starts. The heavy overexcitationof the generators, subject to relief by any necessary engine torque-limiting action, assures the production of maximum system torque and the motor accelerates to near-synchronous speed as an induction motor. The FIELD lever is now moved to position No. 2 and the motor synchronizes with the generators when its field is energized. This motor synchronization can readily be made automatic under the action of a slip sensing relay, which not only makes certain that the motor speed is sufficiently close to assure successful synchronization, but signals the application of its field a t the most favorable instant of its slip cycle. The final movement of the FIELD lever to the RUN position latches the motor field contactor in the closed position. The propulsion regulator now will maintain the bus voltage a t a constant volts-per-cycle and the machines will operate close to unity power factor. The SPEED lever can now be moved to any speed in the SLOW to FULL range with the entire system (engine generators and main motor) moving together in synchronism. The starting and running characteristics of the system are shown in Fig. 21 for six engine-generator sets in service and for three sets in service. Since typical diesels cannot be depended upon for reliable operation a t less than about 30-percent speed, the minimum propeller speed is also 30 percent under normal synchronous generator-motor operations. From a practical standpoint this is a satisfactory condition in almost every application, but if some lower speed must be provided, a subsynchronous mode can be utilized. The main motor, operating as an induction motor, is run a t high percentage slip by controlling the generator excitation to permit such action. Operation in this mode is within the capabilities of the electrical equipment because the propeller power demands below 30percent speed are very low. Once the ship is moving at relatively high speed (above 60-percent speed or so) considerably higher

ELECTRIC PROPULSION DRIVES

torques are encountered in reversing the propeller and stopping the ship. A reversal from high ahead speeds requires that special steps be taken to slow the vessel down part way before the engines can assume the burden within their capabilities. The situation can be seen from an examination of Fig. 22. The motor torque developed under the system constraints is unable to reverse the propeller until the ship headway has been reduced to about 65 percent. If the motor were unwisely put into such a process, it would be subjected to a long period of heavy plugging action and excessive heating of its amortisseur winding would result. A much more satisfactory situation results from the use of a dynamic braking process during the necessary ship deceleration period and the subsequent use of the motor in its induction motor mode. A dynamic brake is a resistor which can be connected to the motor terminals to absorb power generated in the motor by the action of the water passing through the propeller due to the ship's headway. This action does not produce heat in the amortisseur winding, and it is a highly effective method for developing astern thrust, particularly in the upper and more critical high-speed range. When the ship's velocity has been reduced to the necessary 65 percent, the dynamic brake can be disconnected and the propulsion motor connected to the propulsion bus in the reverse phase rotation. The motor can now be plugged and accelerated in the reverse rotation within the capabilities of the engines and synchronized when it gets close to the 40-percent speed of the generator sets. After synchronization, the motor speed can be increased quite rapidly and the vessel stopping procedure completed. This process is illustrated by Fig. 23. From a control sequence standpoint, the foregoing reversing actions are carried out from a full AHEAD condition in the following manner: (1) Return the SPEED lever to the MANEUVERING position. This positions the engine governors at 45-percent no-load speed. However, the complete system continues to run at 65-percent speed, due to the forward movement of the ship and resulting water action on the propeller. (2) Return the FIELD lever to the OFF position. This disconnects the propulsion motor field and reduces the field supplied to the generators. (3) Move the REVERSING lever to the DYNAMIC BRAKE position. This connects the propulsion motor to the brake and applies the correct motor field. The control lever is left in this position until the ship's headway has decreased sufficiently to permit the motor to be pulled into synchronism astern. (4) After the REVERSING lever is moved to the ASTERN position, the FIELD lever and SPEED lever are moved in the same fashion as described for starting. The generator field is forced to the maximum permitted by the propulsion regulator, the motor is automatically synchronized as it accelerates to near-synchronous speed

Fig. 22

Revenal characbristics

70RPM OR

% SPEED

MOTDR-PROPELLER AND GENERATOR RPM

fig. 23

High-speed rwcmal performance

astern, and the SPEED lever is moved to the desired astern propeller speed. The propeller is accelerated to the desired astern speed at a rate determined by the number of engines in service. The automatic load limit feature of the engine governors limits the fuel, and consequently the torque, to the maximum capacity of the system.

360

MARINE ENGINEERING

Section 6 Electric Couplings 6.1 -General Description. The electric coupling is a device for transmitting torque by means of electromagnetic forces without having any mechanical contact between the driving and driven members [12, 131. The electric coupling consists of two steel spiders with rims and flanges. The inside of the outer rim carries a number of poles which can be excited from an outside source through collector rings. The inner element, a laminated core, surrounds the rim and carries a squirrel-cage winding similar to that of the usual squirrel-cage induction motor. Both the inner and the outer elements are supported for rotation and separated radially by an air gap of about % in. One element, usually the inner one, will be connected to the prime mover, and the other to the driven device. Coupling ratings are usually in the 1000 to 4000-hp range. A typical coupling is illustrated in Fig. 24. The fundamental principle of the electric coupling is that of developing torque by inducing current in a squirrel-cage induction-motor-type winding by rotating a magnetic field around the squirrel-cage winding. The coupling-driven element rotates in the same direction as a

!I

Rg. 24

the driving element but a t a slightly slower speed, and the amount of this "slip" is just that required for the development of the necessary driving torque. I n order for electric couplings to be suitable for maneuvering a ship, they must be capable of producing large amounts of torque a t high slip. Thus they are normally equipped with double-layer, bar-type, squirrelcage windings and are designed to produce 150-percent pull-out torque, as well as a minimum of 75-percent normal torque up to 140-percent slip. Such high values of slip will be encountered during reversal duty when the prime mover is operating in the reverse rotation and the propeller is yet operating in the ahead rotation. The double-layer, squirrel-cage winding enables the coupling to produce high torque a t high slip conditions such as occur during maneuvering operations. When the slip is high, the induced voltage is of high frequency, causing the higher reactance deep bars to force the current into the outer high-resistance bars. This results in maximum torque. When operating at normal slip the frequency is low, and a major portion of the current flows in the .deeper, low-resistance winding,

Electric coupling

ELECTRIC PROPULSION DRIVES

resulting in high efficiency. Efficiency is usually above on Shipboard," IEEE Publication No. 45, February 97 percent, including excitation loss of about 1 percent, 1967. slip of 1% percent, and some windage loss. 2 J. A. Wasmund, "Series- Versus Parallel-Con6.2 Applications. A number of installations of nected Generators for Multiple-Engine D-C Dieselelectric couplings have been made on cargo vessels and Electric Ship-Propulsion Systems," Trans. AIEE, 1954. on large tugboats. The typical arrangement utilizes 3 W. E. Jacobsen, "Marine Power Applications," a pair of moderate-speed reversible engines to drive a Standard Handbook for Electrical Engineers, 10th edition, single propeller via electric couplings and a combining McGraw-Hill, New York. reduction gear. Here the couplings provide for the 4 J. A. Wasmund, "How Trials Prove Design engagement or disengagement of the engines simply by Theories," Marine Engineering/Log, August 1956. energizing or deenergizing the field winding. 5 "CS Long Lines," Marine EngineeTinglLog, July I n the case of a twin-engined ship, the usual procedure 1IYVV. OR9 when maneuvering in very close quarters is to run one 6 D. W. Drews, "Turbine-Electric Propulsion for engine ahead and the other engine astern. The ship may Ships," ASNE Journal, August 1963. then be maneuvered readily in either direction by 7 M. Mack Earle, "The Conversion of T2 Tankers operating a lever which applies field to the proper coupling, and thus connects the propeller to either the for Great Lakes and Seaway Service," Trans. SNAME, ahead-running or astern-running engine. All ahead and vol. 68, 1960. 8 "Sun-Built T2-SE-A1 Tankers, " Marine Engineerastern thrusts, within the capability of one engine, are ing and Shipping Review, July 1947. then attainable without further reversing of the engines. 9 T. W.Bunyan,P.D.Morris,andD.D.Stephen, Electric couplings act as torsionally flexible members and torsional dampers. The pulsations in torque from "Canberra," Trans. IME, October 1962. 10 W. E. Jacobsen and R. L. Koch, "Diesel-Electric the engines are smoothed out, reducing gear wear and Propulsion for Polaris Submarine Tender," ASNE noise and minimizing v torsional vibrations in the drive ~ou~rnal, August 1962. system. 11 H. M. Burford, R. L. Koch, and J. D. Westbrook, The propulsion control system, in addition to the usual engine starting, stopping, and reversing features, requires "Performance of a Diesel Electric A.C. Propulsion Plant only the integrated arrangement of engine-speed govern- (Based on the Design and Sea Trials of USS Hunley ing and coupling excitation control. It can be arranged (AS-31))," SNAME Hampton Roads Section, October to suit virtually any particular requirements and is well 1962. suited to remote or pilothouse control. 12 M. R. Lory, L. A. Kilgore, and R. A. Baudry, "Electric Couplings," Trans. AIEE, August 1940. References 13 M. R. Lory, "Electric Couplings for Great Lakes 1 "Recommended Practice for Electric Installations Ships," SNAME Great Lakes Section, September 1950.

1 I

PROPELLERS, SHAFTING, AND SHAFTING SYSTEM VIBRATION ANALYSIS

I

CHAPTER XI I*

C. L. Long -

f

II System Propellers, Shafting, and Shafting Vibration Analysis

Section 1 Introduction 1.1 General. A main propulsion shafting system transom sterns. The distinguishing characteristic of (including the propeller) consists of the equipment neces- this arrangement is that the shafting must be extended sary to convert the rotative power output of the main outboard for a considerable distance in order to provide propulsion engines into thrust horsepower, suitable for adequate clearance between the propeller and the hull. propelling the ship, and the means to impart this thrust One or more strut bearings are required to support the to the ship's hull. In the following pages, the design of outboard shafting. a main propulsion shafting system will be discussed from A shafting arrangement typical of single-screw merthe viewpoint of a shipbuilder undertaking the task of chant ships is shown in Fig. 2. The arrangement illuspreparing a detailed design. I t will, however, be as- trated corresponds to the so-called Mariner or clear-water sumed that the propeller hydrodynamic design has been stern design (there being no lower rudder support); developed; the hydrodynamic design of propellers and Powever, the shafting arrangements of most merchant other propulsion devices is thoroughly covered in Prin- ships are very similar. The major difference between the ciples of Naval Architecture [I]' and therefore will not be shafting arrangements of various merchant ships is the pursued here. Although the fundamentals outlined in location of the main engines. When the main engines the following sections apply to all types of propulsors are located well aft, such as on tankers, there may be as andc prime movers,. the discussion has been primarily few as one or even no inboard bearings at all. When the directed towards a conventional arrangement with a main engines are located approximately amidships, as fixedLpitch propeller and a geared steam turbine main on dry cargo ships, a considerable length of inboard engine. This was necessaw in order to reduce the range shafting is required. of variations which had to be considered. The shafting located inside the ship is termed line Due to the nonuniform wake field in which a ship's shafting. The outboard sections of shafting (wet shaftpropeller operates, the propeller is a source of potentially ing) are designated differently depending upon their dangerous vibratory excitations. The shafting system location. The section to which the propeller is secured itself, which is inherently flexible, is extremely vulnerable is the"propel1er shaft or tail shaft. The section passing to these vibratory excitations; consequently, an analysis through the stern tube is the stern tube shaft unless the of the dynamic characteristics of a shafting system is an propeller is supported by it (as is the case with most integral aspect of the design process and is discussed in merchant ships) in which case it is designated as the this chapter. propeller shaft or tail shaft. If there is a section of 1.2 Description of Shaftfng System. The main pro- shafting between the propeller and stern tube shafts, it wpulsion shafting system must accomplidh' a number would be referred to as an intermediate shaft. of objectives which are vital to the ship's operation. Shafting sections are connected by means of bolted These objectives are: (a) transmit the power output from flange couplings. The coupling flanges are normally the main engines to the propulsor; (b) support the pro- forged integrally with the shafting section; however, pulsor; (c) transmit the thrust developed by the propulsor when required by the arrangement (e.g., stern tube to the ship's hull; (d) safely withstand transient operating shafts which require flanges on both ends and also require loads (e.g., high-speed maneuvers, quick reversals); (e) corrosion-resistant sleeves to be fitted to the shaft in way be free of deleterious modes of vibration; and df) provide of bearings), a removable coupling, sometimes referred to reliable operation throughout the operating range. as a muff coupling, is used. Figure 1 is a shafting arrangement typical of those Bearings are used to support the shafting in essentially found on multishaft ships and single-shaft ships having a straight line between the main propulsion engine and the desired location of the propeller. Bearings inside the ship are known by several names with line shaft bearings, steady bearings, and spring bearings being the most Numbers in brackets designate Reference8 at end of chapter.

I

iI t

MARINE ENGINEERING STATE PERFORMANCE REQUIREMENTS ESTABLISH DESIGN CRITERIA DEVELOP SHAFTING ARRANGEMENT

ESTABLISH SIZES

DETERMINE DYNAMIC CHARACTERISTICS DEVELOP DESIGN DETAILS

DESIGN COMPLETE

Fig. 3

Shafting system design sequence

popular in that order. Bearings which support outboard sections of shafting are called stern tube bearings if they are located in the stern tube and strut bearings when located in struts. Outboard bearings may be lubricated by either seawater or oil; high-quality seals are required in the event the latter is used. I n order to control flooding, in the event of a casualty, bulkhead stuffing boxes are installed where the shafting passes through bulkheads. A more substantial seal is installed a t the forward end of the stern tube where the shafting penetrates the watertight boundary of the hull. The propeller thrust is transmitted to the hull by means of a main thrust bearing. The main thrust bearing may be located either forward or aft of the slow-speed gear. If located forward, the thrust collar is detachable so as to permit the ii~tallationof the gear on the shaft and, secondarily, to permit replacement of the thrust collar if ever required. If located aft, the collar may be forged integrally with either the slow-speed gear shaft or a subsequent section of shafting. Since one purpose of the main thrust bearing is to limit movement of the slow-speed gear, the main thrust bearing is usually installed close to the gear. Installation of the thrust bearing close to the gear also facilitates lubrication of the thrust bearing. 1.8 Design Sequence. The design of a shafting system is, by necessity, an iterative process because the various system design parameters are, to some extent, mutually dependent. The iterativedesign process usually followed is illustrated in Fig. 3. As indicated by Fig. 3, the first step in the design of a shafting system is to state the performance requirements;

that is, the type of propul$ve system, number of shafts, type of service, and the like. Next, the design criteria to be employed must be fixed; i.e., one of the various classification society rules could be followed, oil-lubricated stem tube bearings may be selected, hollow shafting may be ruled out, etc. I n establishing the design criteria, it must be recognized that the shafting interfaces with thepropulsor, the main engines, and the ship system as a whole. After the design criteria are established and the general ship arrangement is available, an approximate shafting arrangement can be developed. This entails a t least tentatively locating the main engine, propeller, and shaft bearings with due regard given to arrangement restrictions, clearances required, shaft rake, construction restraints, and overhaul and maintenance requirements. Before the design can progress further, the shafting diameters, corresponding to the preliminary arrangement, must be computed along with the length of shafting sections, flange dimensions, and preliminary propeller data. With this data the bearing reactions can be approximated and the bearing dimensions and loadings can be checked. At this point, it will generally be desirable to adjust the bearing arrangement tentatively selected so as to obtain more equal bearing reactions or to alter the number of bearings. Variations in bearing loads due to thermal expansion of the shafting bearings, particularly those in the way of the main engines, are investigated to ensure satisfactory bearing performance under all operating conditions. There are three basic types of vibration which can occur in a main propulsioq shafting system; these are torsional, longitudinal, and whirling vibration. It is essential that a preliminary vibration analysis of the shafting system be made in the early design stages because the shafting vibration characteristics are largely established by the ship parameters that are fixed a t that time. Specifically, the shape of the hull afterbody, type of propeller, propeller aperture clearance^, number of propeller blades, length of shafting, shaft material, position of the m+in thrust bearing, type and configuration of prime mover, spacing of the aftermost bearings, and type of aftermost bearings largely establish the dynamic characteristics of a shafting system. The subsdquent development of design details has a relatively secondary effect as compared with these major parameters. In addition, an analysis of the system's response to shock loadings is required for naval combatant ships. An analysis of the dynamic characteristics of a shafting system can be one of the more complex aspects of the design process. Once the arrangement, component sizes, and dynamic characteristics have been shown to comply with the d e sign criteria, design details are developed. This entails designing flange fillets, flange bolts, keys, keyways, sleeves, and the like.

PROPELLERS,SHAFTING, AND SHAFTING SYSTEM VIBRATION ANALYSIS

Section 2 2.1 Location of Main Engines. The engine flange location and the propeller location are essential information in establishing the shafting arrangement. The foreand-aft position of \he main engines is generally established during the preliminary design stages after studying the ship cargo stowage, ship trim, and shafting system. The cost and n~eightof shafting are significant; for these reasons, and also to minimize the use of prime cargo space for the main machinery and shafting, the main machinery is located as far aft as practicable. With vessels such as oil tankers, the main machinery is confined to the stern end of the s h i ~such that a short run of shafting is required and the ;umber of line shaft bearings is minimal. On the other hand, dry cargo vessels do not have the ability to adjust their operating draft by taking on ballast; therefore, in order to provide satisfactory light load draft conditions, it is necessary to locate the main engines (and associated weight) well forward of the stern. Normally the main engine should be set as close to the inner bottom as the configuration of the main machinery will permit. It is possible, and it is the usual case, to have limited projections of the main machinery (e.g., the slow-speed gear lube oil suqp) below the inner bottom when such projections do not excessively weaken the inner bottom (see Figs. 6, 8, and 9 of Chapter I). The main engine location in the athwartship direction is on the ship centerline of single-screw ships. On multiscrew ships the engines are set off the ship centerline approximately the same distance as the propellers, but the shaft centerlines usually do not parallel the centerline of the ship. The location of the engine in the athwartship direction is controlled by the propeller location, main engine details, and the machinery room arrangement requirements. 2.2 Location of Propeller. The location of the propeller is determined by the propeller diameter, the acceptable clearance between the propeller and the baseline of the ship, and the acceptable clearances b e tween the propeller and the hull in the plane of the propeller. Although the propeller diameter selected should theoretically be the one corresponding to optimum efficiency for the propeller-ship system, in practice the optimum propeller diameter is usually larger than can be accommodated. As a result, the propeller diameter selected is a compromise. In locating the propeller in the aperture of a singlescrew ship, a clearance of 6 to 12 in. is normally provided between the propeller tip and the baseline with clearwater sterns-or to the rudder shoe with a closed stern (Fig. 4). With high-speed ships, which are generally characterized by shallow draft and multiple screws, propellers are often permitted to project below the base line in order to provide adequate clearance between the propeller and the hull. This is satisfactory provided

maxiinum draft limitations for service routes or drvdocking are not exceeded. One of the most effective means of ensuring a satisfactory level of vibration aboard ship is by providing adequitte clearance between the propeller and the hull . surface. For this reason, the subject of providing clearances is one of overriding importance. Generally speaking, the greater the clearances, the better the performance from a vibration standpoint. There are three types of vibratory forces generated by the propeller: (a) alternating pressure forces on the hull due to the alternating hydrodynamic pressure fields caused by the propeller blades; (b) alternating propeller shaft bearing forces which are primarily caused by wake irregularities; and (c) alternating forces transmitted throughout the shafting system which are primarily caused by wake irregularities. If the frequency of the exciting force should coincide with one of the hull or shafting system natural frequencies, very objectionable vibration can occur. A further breakdown of the forces generated by the propeller is given in reference [2]. When selecting propeller clearances, the perfprmance of similar ships should be an influencing factor. Of course, differences between the important parameters of the ships under comparison must be assessed. Important parameters to consider are the unit thrust loading on the propeller blades, number of propeller blades, amount of propeller skew, length of the ship, and t& ending angle of the water-plane forward of the propellet. References [3,4] discuss the influence of the more important of the foregoing parameters and summarize test and analytical data on this subject. Figure 4 may be used as guidance in assessing the aperture clearances of single-screw ships. Figure 4 shows the

Fig. 4

Propeller aperture clearances

. MARINE ENGINEERING athwai-tship rake angle, both ~f which are measured relative to the ship centerline, It is rare for 19 to exceed 3.75 deg or 4 to exceed 2.5 deg. From rake alone the reduction in propulsion efficiency will normally not exceed0.3pei-cent. Aside from the efficiency penalty, there is no objection to moderate amounts of rake. 2.4 Shaft Withdrawal. Occasionally shafting sections, particularly those outboard, must be withdrawn to be inspected or repaired. Consequently, provisions for removing shaft sections from the ship must be considered when developing a shafting brrangement. . On singlescrew ships with shafting arrangements similar to Fig. 2, the propeller shaft is almost without exception withdrawn inboard for inspection. If repairs are necessary, the shaft is removed from the ship by cutting a hole in the side of the ship and passing the shaft through it. This technique u.ould be used for removing line shaft sections as i-ell. Ftfi. 5 Ueamnm of a propelk supported by strut bearing With shafts having struts as shown in Fig. 1, a check ahaft can be must be'made to ensure that the withdrawn from the strut after the propeller is removed. range of eexperience wIdah hw been obtained in cannec- Withdrawal can be accomplished by removing the beartion dth I a q p single-screw &ips. When the propeller ing bushings so that the shaft can be inclined sufficiently is supported by a strut bearing, i.e., multiscrew and to aiIIow the forward end of the shaft to clear the ship's bmam-&ern vesseh, two clearance dimensions 11-amnt structure, mating shaft flange; etc. This consideration careful skudy. These dimensions rand the range of ex- can govern the length of the propeller shaft and the size of the stmt barrel. Figure 1 shotvs the removal position perience with them are shown in 1%. 5. 2.3 Shaft Rake. In order to provide ~atiiudein of the prcrpeller shaft., Removal of the stern tube shaft, which must hhve locating the position of the pmpeller and the main engines, it is usually necessary to wke the shaft cehterline. flanges on both ends, requires a decision regarding the The &aft is generally raked downward going aft as this type of flanges to be provided on the shaft. If the shaft permits the main engines to be located higher in the is manufactured with integral flanges on both ends, the ship. In mdtiscren- ships the shaft is generally raked in stern tdbe barrel and bearing bushings inust be sized to both the ve$ical and hori~onG1planes, usually donm- pass the flange diameter. Since i t is desirable to pass the shaft outboard, sufficient clearance should be proward and autbalbrd going aft. Large rakes s h a W be a~oidedsince a reduction in the vided to incline the shaft such that it will clear outboard propulsive efficiency is associated with rake. The intm- struts, etc. I n order to use smaller stern tubes and duction of rake incurs a reduction in the propulsion M a g bushings, the stern tube shaft can be manufactured with a removable Aange mupling on the forward efficiency equal t o erid. Prior to unshipping the shaft, the removable coupling is removed so that it is not neeesary t o disturb where B k. the shaft vertical rake angh and is the the stern tube bearings. (I

PROPELLRZS, SHAFTiNG, AND SHAFTiNG SYSTW VlSRAMON ANALYSIS

dimtly obtained from the min engine t o q u e and the p r o p e k thrust. On the other hand, vibrabry loads emanating from the propeller do not lend t h m m i ' ~ to s a precise evalwtion and are diieuit tn tseat in an absolute sense. 3.2 P r o p e l l d n d u d Loads. Aside from the alternating bending s h due to Ithe weight of the propeller, the ckumferentially n o n d o r m velocity of the water inflow to the propeller (wake) is the most important qouni.e of the alternating i d s in the shafting system. If is, howeverS important to $i&hguiish between the importance of the chumfereniaa nonuniforrnity of water i d o w a t a particular propeller radius and the nonunifomity of the average flow at, one radius as compared with mother. W e the former 1eads to vibratory propeller fothe latter does @. A propeller blade section w o r e in a constant d o c i t y field & a particular radius has a steady flow and force p-hkrni. The average axial velocity at each radius can be dierent without @awingalternating loads. I n such a ease the pmpeHq design can be adjusted for radial variations in the inflow vdocity to aehieve optimum efficiency. However, a propeller can only be designed t o satisfy average eond%ions a t each radius. Variation in the axial component of the inflow velocity a t each radius gives rise to the p e r i d c fluctuating forces genembd by the propeller. The variation in the inflow water =loci@ at a particular radius results in a change in the angle d attaek of the psopeI.3~blade sections as the propeller makes one ~vo1ution,thereby creating a1krnati.g propeller forces. Figure 6 b an example of the axial, VA,and tangentid, VT,i d o w velocities in the plane of the propeller for a single-screw ship. The tan- . gential velocity component is symmetric on both sides of the vertical mnterhe and is g&erally upward. The symmetry of the tangenttd ve1ociQ component would

+

rw VA V,

tend to suggest that its effect is uniform, but such is not the case. For a propeller bhde rotating clockwise looking forward, the tangedtiaJ veJoci*y component, eRwtively reduces the angle of attack on the bhde sections as they pass up the port side (reducing thrust) and increases the angle of attack of the blade sections as they pass down the starboard side ( i n c b n g thrust). Figure 7 iliustmtes how the variable axial and tangential velocities give rise t o variable loads. Also, another very important fact is t h t the tangential velocity components shift the center of propeller thrust to the shrboard side of the

speed of advance

= resultant of V, and V, = minimum advance angle ,P = maximum advance angle E . = variation in blade advance Ya

Section 3 f ftafting Leads 3.1 Dwkn Coaaidemt?ons. In general, the dimensions ~f s h d t b g are predic@ed on the basis of strength requirements; however, it is ocxasionally nebee+ eary to modify an otherwise sstipfactory shafting system design due to vibration considerations. Shafting &rametemurnally have only a minor impact on the longitudind vibration chmcteri~tim~ but the wh&:ling and torsional modes rtre sensitive t o &a& diameters. Shafting vibration, m mch, is d k u d in Sections 7-9.

Propulsion shafting is subjected to a variety of steady and alternating Ioads which induce torsional shear, axial thrust, and bending stressesin the shafting. In addition, there are radial compressive stressesbetween the shafting and mating elements (such as the propeller and sleeves) which, when coupled with axid strains from bending stress, tape v q important from a fatigue standpoint. The steady loads represent average conditions ~nnd ean be &mated w i a~degree of certainty as they are

P,,

Fig. 7 Typical warMan in udvdvence onale of a Made sedan &kg one revdution

Fspd ~ l o w o f w u t c r i n p ~ a n c o f p r o ~

blade tangential veloci'ty = axial speed of advance

= tangential

367

angle during one revolution

Locus of the resultant of tspeeds h e axirri tangential of and odvonce

I(

368

PROPELLERS, SHAFTING, AND SHAFTING SYSTEM VIBRA'I'ION ANALYSIS

MARINE ENGINEERING

much greater for the five-bladed propeller than for the Table 1 Ratio of Shaft Torque Measured During High-Speed four or six-bladed propeller. For a single-screw s h i ~ Maneuvers to Normal Torque having a propeller with an even number i f blades, thk No. Torque Ratio fluctuating forces of two opposite blades give rise to a hi^^^^^ Shaft Inboard Outboard larger total t h r u ~ tand torque amplitude because op- ~~~~l 4 1.2-1.4 1.2-1.3 posite blades simultaneously pass through the slow water Naval 1 1.1-1.2 velocities at the top and bottom of the propeller disk. 2 1.2-1.3 1 1.1-1.3 The transverse force and bending moment,developed by one blade tend to be compensated by similar loads on the opposite blade. For propellers having an odd number of blades, the Table 2 Propeller Variable Torque Excitation Factors blades pass the upper and lower high-wake regions alternately. The total thrust and torque variations are No. of Torque Excitation Factor, r Propeller Blades 3 4 5 therefore smaller as compared with a propeller having an 0.07-0.12 0.10-0.15 0.06-0.10 even number of blades. However, due to the alternate Single-screw vessels Twin-screw vessels 0.02-0.05 0.02-0.05 0.02-0.04 loading of the propeller blades, the transverse forces and with struts bending moment do not cancel. Therefore, larger bend- Twin-screw vessels 0.04-0.08 0.04-0.06 0.04-0.05 with bossings ing moments occur with a propeller having an odd Note: Excitation torque = 4, where Q = mean torque. number of blades. The nonuniform character of the water inflow to the propeller can be resolved into Fourier components with the propeller rotational frequency (shaft frequency) as the fundamental [10]. Since it may be assumed that in power; this results in a higher shaft torque. As the linearity exists between inflow velocity variations and hull becomes foul, the ship speed reduces and full power propeller blade force variations, the Fourier components is developed a t a lower rpm; consequently, the torque of the inflow velocity are also the Fourier components loading on the shafting correspondingly increases. Such of force of a single blade making one revolution. Only torque increases are normally not considered in merchant those harmonics of loading which are integral multiples practice because merchant ships do not engage in extenof blade frequency ( M )contribute to the unsteady thrust sive high-speed maneuvers. The torque increase (which and torque, and only those harmonics of loading adjacent is relatively small) due to hull fouling is accepted as a to multiples of blade frequency (kZ f 1) contribute to reduction in the factor of safety. the unsteady transverse forces and bending moments [5]. The torque increases measured during trials of singleAll other harmonics of shaft frequency cancel when screw and multiscrew ships in high-speed turns are given summed over the blades. The selection of the number in Table 1. The torque ratio shown is the peak torque of blades can be based on the relative strengths of the value observed during steering maneuvers divided by harmonics in the inflow water velocity to the propeller the torque a t the start of the tests. to minimize the alternating thrust and torque and bendAlternating torsional loads on the shafting are gening moments. erated by the propeller and occur at predominantly blade Variable propeller forces, in addition to those resulting frequency as a result of the wake as discussed in Secfrom a nonuniform water inflow, are generated as a result tion 3.2. Although alternating loads can be generated of the proximity of the hull to the propeller. Hull by other sources, the propeller is the only one of practical surface forces generated by the propeller are of the ut- importance, except in diesel propulsion plants, where the most importance when evaluating hull vibrations. cyclic engine torque is significant. Shafting systems are 3.3 Torsional Loads. The torsional load. on the carefully designed to avoid torsional resonant frequencies shafting, which results in the steady torsional stress, is a t full power; therefore, alternating torsional loads are calculated from the output of the main engine. If the not congidered to be amplified by resonance. The range full-power shaft horsepower output, shp, of the main of magnitude of the forced torsional alternating loads is engine is developed a t N rpm, then the steady torsional given in Table 2. I t will be noted that the variable load, Q, on the shafting is: torque can be of a significant magnitude even without magnification. 3.4 Thrust Loads. The magnitude of the steady thrust load on the shafting system is equal to the towed I n the design of naval shafting systems, it is common resistance of the ship a t the speed corresponding to maxipractice to increase the torque calcu!ated with equation mum design power, corrected by the interacting effect (2) by 20 percent. The increase in design torque is an of the propeller and hull as the propeller pushes the ship. allowance in recognition of the additional torque de- This interaction effect is known as the thrust deducveloped during high-speed maneuvers, rough-water op- tion [I.]. The value of the design thrust can be obtained erations, foul-hull conditions, etc. During turns, the from the powering calculations or from model basin tests propeller rpm reduces without a corresponding reduction of the ship. For preliminary design purposes

-

.&

I 0'1

E

I

I

90. (a1

180. PROPELLER POSITION

270.

360.

h w

( c ] VERTICAL BENDING MOMENT IN PROPELLER SHAFT (PROPELLER WEIGHT INCLUDED)

$

a w

PROPELLER TORQUE VARIATION

E~:EE

PROPELLER POSITION

z

+

!j I5 J

I

I

3+10 4

K w

p+

5

I

0

I

0.

90.

180' 270. PROPELLER POSITION

360'

( d l HORIZONTAL BENDING MOMENT IN PROPELLER SHAFT

0 I'.

5

0.

0

w

a- 5

i*'

b

,

'b'

L"

I

0.

'v' I

90'

180' 270' PROPELLER POSITION

360.

(b) PROPELLER THRUST VARIATION

B

Fig. 8

I I

i/

Typical single-screw propeller alternating thrust, torque, and bending moments from nonuniform water inflow velocities

propeller centerline of a clockwise-turning propeller on a of the propeller, the thrust and torque can be determined single-screw, ship. This off-center thrust gives rise to a and plotted as shown in Fig. 8. If the K r K e - J diagram bending moment which is imposed upon the propeller is not available for the propeller, the step-by-step calculation in reference [7] can be used. This method is shaft. Analyses can be made to predict the magnitude of the based on that given by Burrill in [8]. Since the slowest axial inflow velocity (highest wake) alternating components of torque and thrust including the eccentricity of the resultant thrust relative to the of single-screw ships is generally in the region above the shaft centerline [5]. Four basic methods are available to propeller centerline, the greatest thrust tends to be calculate the unsteady forces and moments on marine developed when the propeller blade is in the upper part propellers aaused by circumferential nonuniform inflow. of its orbit. The effect of the tangential inflow velocity These are quasi-steady, two-dimensional unsteady along is to shift the resultant thrust to the starboard side bea strip, combination quasi-steady two-dimensional un- cause the propeller blades develop greater thrust moving steady along a strip, and three-dimensional unsteady. against the tangential velocity, as discussed in the foreEither the quasi-steady or the two-dimensional unsteady going. This subject is given a detailed discussion in technique may be used to obtain approximate estimates reference [9], and it is noted that as the shape of the stern of the fluctuating thrust and torque; but if close predic- sections change from a V to a U shape, the resultant tions are required the three-dimensional unsteady tech- thrust center tends to move down because the inflow nique should be used. Application of the quasi-steady velocities over the bottom region of the propeller disk method is much simpler than the three-dimensional un- become more nearly equal to those in the upper region. steady approach. Due to its simplicity and the fact that The position of the resultant thrust is also sensitive to it produces results which are generally accurate enough the ship's draft. For instance, when a cargo ship operin a relative sense for most practical applications, the ates lightly loaded with the propeller blades breaking the water surface, the center of thrust obviously shifts quasi-steady approach is a very useful method. A quasi-steady analysis is conducted by making an lower in the propeller disk. Figure 8 shows that a single-screw ship with a four or instantaneous examination of the flow velocities relative to the propeller blades a t discrete angular positions of a six-bladed propeller (that is, an even number of blades) propeller blade [6]. The inflow velocities are regarded has larger torque and axial thrust variations than one as constant (quasi-steady) at each blade position. By with a five-bladed propeller. However, the thrust eccenusing the open-water characteristics (KT&-J diagram) tricity (propeller shaft bending moment) is shown to be

1 t

,'

369

-

PROPELLERS, SHAFI'ING, AND SHAFTING SYSTEM VIBRATION ANALYSIS

Table S

Propdler Variable Ruwst Exc8atiorr F a c m That EHcitation No: Bk&s ~EEEW~-&~W k B h Frotar, J Behind struta O.W.85 %%or& Behind ekege 0 . ~ 4I% . 3 4ar5 w i d akee 0.05-0.W R Behind hossinea O.WU.12

Behind b & i

Behind bmaing~t

0.w.10

Q.W.08

Q.OEHl.12 a U-seotine tend to emphsiae the e v e m d e r compomntg and B-wotiona the & d e r eamponents. Nate: M a t i o n thrask = p,where T = m e n thru&. s

sional the imps-ce of the k t stis reduced even furthepep 3.5 B e d L d s . Lmds which cause bending &mes to omur in the shaP%iqm e the d t of gravity, shmk, off-center t M loarfsBand whirling shaft vibmtion. With the exceptionof once-per-revolution whirling vibration. d are &ernratinn hack relative to a point on t i e shaft'a~~cl occur at e&h& shaft rotative fre&eney or Qnce or twim propeller blade frequency. The weight of the shafting iW (a gravity l o d ) * k

are

shaftini unless there =US& w*ht ~ ~ ~ l ~ e f t t m t i ~ ~ , such as a S h d t locking device or brake d r m , or an exceptionably long span between bearings. When the shaft mans beheen bearilarrs are essentially equal, the maximbrn static bending moment a e c t m b ~ aihe t shaft bearings as a result of shaft weight cae be d e f e r m i d

Table 4

Ship Name OWsp Chrysai Robimon Jamestwn Observaiion zslund

II

heh hue

of f mnges from about 0.16 to 0.23 for singIescrew &hipsv e n g frarn fine to full lines, respectively. Twin-screw ships have t values ran& from about 0.1 t o 0.2,2,& larger value applying t o &ips with bossings and tihe sm&w value mmspon&ng to &ips with struts. PC values of 0.73 for Singlewrew ships and 0.68 for mdtiscrew ships sre average v a b and normally found to be suitable for preliminary estimates. Reference tII contains methods of estimating t and PC and should be consulted €0 obtain a more a m r a t e &mate of the thrust I d if m d d test results a~ ship's performance etiYenleiti0~~ are n ~ avaihble. t Alternating thrust h a & ltre generated by the propeller. The p d o m i n m t dtemating t h r a 1-d o m m at pmpdler bMe frequency abi a consequence of the nonunifwm inflow water velacity 1;6 the propeller as discussed in Election 3.2, The magnitude af the variable thrusC M a is dependent an the ntmbes of prnpeller blade& For single-mrew ships, an even number d bJr8des will m l t in greater d t e m t i q thrust lmds than am odd n u m k as dimussed in 3.2. For preEminmy e&imates, fhe magnitude of the alternating thrust as a percentage of steady thrust can be taken from Table 3. Insofar tm the strength of the shafting is concernedr neither the steady nor J k m a t i n g thrust lmds are major design consi~ations. With merchant ships, the steady camp&ve stress is 1008 to I500 @; even In highly stressed shafts in navd skips the steady e o m p d v e stress d d o m reshes 2500 psi, Torsional shear stresses me of predominant importance; and since the s t m s due to thrust EEO nut combine additively with the tar-

Tanker Mariner

20,300 38,100 15,200 50,200

16,400

2

M,

=

Ib/h.

@ are mt approxirnateIy If the spans bekeen equal, such a simple appmmh cannot be used; instead, continuously supported beam analytieal techniques, such a5 the three-moment equation or Hardy C m methad, must be used. During recent y e w , the practice has been to use the digital mmput;er to mIculate the bending moments a t all efiticaI &a& sections, utilizing mntinuous b a r n forrnuIa%ians. Weight 1 4 s on the outboard shafting tend to be of more imporbnee due to the large mncentraW weight of the propeller. Standard eonfirnous beam equations ean be used ha determine the magdude of the bending moments. Howeuw, because of the long bearings used o u t b d , the lacakians of fhe bearing peaetions we not cleady defined. The pactice is ta assume the reaetion a t the e n t e r of d1 bearings except the bearing just fopward of the propder. Because af the ham weight of the propelIer, the pmpeller shaft has a s i ~ i 6 c a n slope t a t this bearing; therefore, the rwltanh W i n g reaction tends to be in the after region af the Wing. WaterIubricated bearings of Emurn vitae, micarby or rubber have L / D ratios of about 4 f o r this bearing and the resultant reaction is usually assumed to be one shdk diameter forward of the aft bearing face. Od-Eubrieated bearings h&veLID ratios of about I to 2, and a review of the shaft contact in these bearings indicates that hard contact is confined to the after region of the bearing for a length appmximateIy eq;uaI to the diameter of the shaft. Current practice is t o assume that the resultant bearing reaction in oil-lubricated b e d g a Is one-half shaft diameter from the after bearing face. Generafly the most signifimnt weight moment ia due to the overhtmg moment of the propeller. The maxi-

Thrust,

Thrust Eccent. C, it

Thrust

Eccent. Factor C / D Reference 0.047 [11]

shp

rpm

503 68 615 84 436.5 62

0.74 0.77 0.69

10,000 15,000 8,500

95 112

184,000 246,000 156,000

19.5

0.91

22

20.5

1.31 1.67

0.060 0.082

B2] [13]

93 76

0.75 0.60

26,500 22,000

108.5

385,000 271,000

23 22

1.26

0.055 0.076

1141

685

528

WpLp

85

110

Table 5

(5)

where =

Equation (5) is the moment at the bearing reaction point assuming that the reaction is a point support rather than a distributed reaction over a region of the shaft. The point support assumption is justified in that the exact load distribution on the bearing is unknown and the moment calculated in this manner is somewhat in excess of the actual value when the position of the resultant reaction is estimated reasonably well. There are a number of influences in addition to the gravity moment of the propeller which can have a significant impact on the propeller shaft bending stress. These are the eccentricity of thrust, water depth, sea conditions, and ship maneuvers. Under the general guidance of SNAME Panel M-8, the propeller shafts of a total of five ships have been instrumented to measure the bending stresses under actual operating conditions. Data obtained from these tests are reported in references [ll-151. Table 4 summarizes the characteristics of the ships tested. The tests were conducted to show the significance of the ship loading, sea conditions, ship maneuvers, and thrust eccentricity. Eccentricity of the propeller thrust produces a significant propeller shaft bending moment. The propeller resultant thrust is eccentric from the propeller shaft centerline under almost all operating conditions and is usually in the upper starboard quadrant when looking forward. Therefore, it does not combine directly with the propeller gravity moment. Light draft operating conditions and "U" shaped stern sections tend to bring the thrust and gravity moments closer together and make them more additive. Table 4 shows the thrust eccentricity factor, C / D , determined from full-scale test data for heavy-displacement, calm-sea conditions. The thrust eccentricity, C, shown in Table 4 is the resultant of the eccentric thrust and the gravity components.

ib

Prop. Dia. D, ft

cn

A

propeller overhung moment in propeller shaft, in.-lb W , = weight of propeller assembly including shafting aft of reaction point, Ib L, = distance from CG of propeller assembly to aftermost bearing reaction, in.

M L

= bending moment at bearing*in.-lb = span between in. w = weight per unit length ef shaft&

Rhin ."---=

Lyl

Design

mum static propeller shaft bending stress is computed as

M,

V = ship s p d at maximurn power, knots W = ship's resistance at Tr, Ib ehp = ship's bull &wGm horsepowe~a t TP, hp s h= ~ maximum &aft horsemwer, h~ *t =. thnrst cE&uction fraction PC = propulsive coefficient

Type Ship TZSE-A2 Tanker - -. . --. Tanker Victory

Ships Instrumented to Determine Tailshaft Bending Stresses

Load Heavy Heavy Heavy Light Light Light Heavy Light Heavy Light

1.M

[16]

Increase in Propeller Shaft Bending Stresses Due to Various EfFects

Sea Condition Calm Calm Calm Calm Calm Calm Stormy Stormy Calm Calm

Operation Ahead Ahead Maneuvering Crash Back Ahead Ahead Ahead Ahead

Water D e ~ t h Factor

Deep Deep Deep Deep Shoal Shoal

3% 9%

2%

1% 1%

lhll-scale tests on the Esso Jarnestown I141 permit an evaluation of the influence of ship loading, sea conditions, and maneuvers. These factors are summarized in Table 5. The factors presented in Table 5 are the ratios of the bending stresses for the various conditions described to the bending stresses under full-load, deepwater, calm-seas, and straight-ahead operations. The extrapolated results from the Observation I s l a d tests [15] generally support the factors in Table 5. It should be noted that maneuvers such as era&-backs rarely occur, and that the shaft need not be designed to withstand stresses three times the normal value on a continuing basis. Shock loadings, considered in designing naval shafting for combatant ships, are akin to the gravity loading and are frequently determined by multiplying the gravity force loads by a "shock" factor; however, more sophisticated methods are available for determining the shock loads through the application of dynamic analysis techniques. References [16,17] treat the procedure for conducting dynamic shock analyses of shafting systems. Misalignment in shafting systems can produce very significant bending loads and this factor is probably responsible for the majority of inboard shafting failures. The sensitivity of the shafting to misalignment should be reviewed particularly as regards water-lubricated stern tube and strut bearings which are subject to wear in service. The sensitivity of the shafting to misalignment can be assessed by calculating the shafting bearing reactions and moments with the shafting in various misaligned condi-

372

PROPELLERS, SHAFTING, AND SHAFTING SYSTEM VIBRATION ANALYSIS

MARINE ENGINEERING

I

'

tions. Digital computer programs have been developed for making this analysis; one such pmgram is described in Section 5. Lateral or whirling vibration of the shafting can result in inqreased bending loads in the shafting. However, since the shafting system is designed to avoid whirling criticals in the upper operating range, bending loads from shaft whirling vibration are not considered when designing the shafting. 3.6 Radial Loads. Radial loads in shafting are caused by driving the propeller onto the shaft taper, shrink-fitting sleeves on the shafting, and shrink-fitting removable flange couplings. The radial compressive stresses resulting from these loads are normally of in-

significant magnitude and are not considered in determining the shaft factor of safety. However, these radial loads can be of importance in that they give rise to fretting corrosion when coupled with bending loads and alternating torsional loads that cause minute relative movement of the mating surfaces. Fretting corrosion can be controlled by limiting the relative motion and by cold-rolling the mating shafting surface. Cold-rolling of shafting surfaces is discussed in Section 4.3. Another consideration is that if the radial load is applied abruptly, a stress concentration can occur. Therefore, design details should be developed so as to minimize sudden changes in radial loads caused by shrink or press fits.

Section 4 Shafting Design 4.1 Shaft Materials. With the exception of naval vessels and merchant vessels of very high power, mild steel is used for both inboard and outboard shafting. In the case of high-powered ships, the inboard shafting may be made of high-strength steel; however, high-strength steel is not recommended for outboard applications. Because of the seawater environment and fretting corrosion conditions that exist at shaft sleeves and the propeller interface, the fatigue limit of high-strength steel is not reliably greater than that of mild steel, nor is the endurance limit in a fretting corrosion condition better than that of mild steel. Considerations in the selection of shafting materials are: fatigue characteristics, weldability, the nilductility temperature, and the energy absorption capability. An array of chemistry and physical property standards has been established for marine shafting materials that provides a range from which shafting materials can be selected. Chapter 22 contains more specific information regarding shafting materials. 4.2 Computation of Shaft Diameters. Shafting for merchant vessels is required to meet the minimum standards set by the classification society which classes the vessel. Classification societies use rather simple formulas to compute the minimum shaft diameters. These formulas normally contain coefficients which are changed from time to time in recognition of experience or advancements in technology. The American Bureau of Shipping (ABS) line shaft diameter formula is of the following form [IS]:

M

d = C(K?) where

d

=

minimum line shaft diameter, in.

c = constant for type of shaft = constant relating to operating environment

K

It may be noted that equation (6) neglects bcnding loads, alternating loads, and stress risers. Furthermore, it presupposes that the shift will be a solid forging of mild steel and that no unusual circumstances exist. However, despite the factors not explicitly considered, equation (6) does provide,a sound basis for designing line shafts. This is because the predominant torsional shear stress is properly considered. The level of torsional shear stress corresponding to equation (6) can be determined by observing that

By substituting equation (6) into equation (7) and setting K = 64, which is the case for oceangoing vessels

If c were specified as 0.95 (as was the case for a period before 1965), the corresponding allowable steady torsional shear stress would be 5849 psi; with a c value of 0.875 (which was adopted in 1965), the corresponding torsional stress becomes 7486 psi. When the allowable torsional shear stress of 7486 psi is compared with the material minimum tensile yield stress of 30,000 psi, it is seen that adequate margin is allowed for the secmdary duences which appear to be neglected. Only vessels to be navigated in ice require special consideration. (6) The stern tube shaft diameter required by ABS for merchant ships is directly proportional to the line shaft diameter. The propeller shaft diameter is related to the line shaft diameter and the diameter of the propeller. The ABS propeller shaft diameter formula is of the following form :

T

=

kld

+ -PC

where

T = minimum propeller shaft diameter kl = constant d = required line shaft diameter P = propeller diameter C = constant relating to propeller shaft environment

Although the formula for the tailshaft diameter is simple, it does recognize that bending stresses from propeller weight and off-center thrust exist and relates these important factors directly to the propeller diameter. It has been expressed by some designers that classification rule$ for tailshafts are not adequate because ,the level of bending stress is not controlled [19, 201. A designer should review his own design against the important parameters and against his experience. Classification rules should not be accepted without question, particularly for unusual designs since the formulas are set primarily by past experience. In any case, the propeller shaft is customarily manufactured with a diameter approximately 3 percent larger (10 percent stmnger) than the minimum classification requirements in order to provide an additional margin of safety as well as to make provisions for removing a small amount of surface metal in the event the shafts become superficially damaged. Reference [21] reports on a service life comparison of 15 oversized shafts (74 percent greater section modulus than required by ABS) with 15 shafts of normal size (11.5 percent greater section modulus). The comparison showed that the mean expected service life of the oversized shafts was less than the shafts of normal size. Although the statistical sample was small, the study clearly showed that propeller shaft problems are not necessarily solved by simply making the shaft larger. The approach used to establish the size of naval shafting is considerably different from that used with merchant shafting. The procedure used to determine the size of naval shafting is delineated in reference [22]. As may be noted from the reference, an effort is made to assess all significant shafting loads in each particular case, although some loads are by necessity handled in an approximate manner. For example, in order to allow for the effects of off-center thrust and abnormal loadings due to rough weather and the like, the propeller shaft bending stress due to the static weight of the propeller is multiplied by a factor of 3 for single-screw ships and 2 for multiple-screw ships. An additional difference between merchant and naval procedure is the criteria of acceptance. In naval practice, dual criteria are used. Factors of safety are specified for.al1 shafting and, in addition, a specific bending stress limit is specified for the propeller shaft. The reason for the latter requirement is that fatigue tests run on models of propeller shaft assemblies and crank pins [13] showed that bending stress levels in surface rolled (cold-rolled)

373

shafts in excess of 6000 psi would result in shaft cracks. Therefore, it is not prudent to design with normal operating bending stresses in excess of this stress level. Furthermore, the endurance limit of a propeller shaft assembly can be essentially independent of the fatigue limit of the material in air. If seawater contacts the steel shaft, no endurance limit exists and it is ohly a matter of time before cracks will occur followed by ultimate failure. 4.3 Bearing Locations. In the past, bearing locations have been determined by criteria such as "each shaft span shall have two bearings" or by intuitive judgment. With these criL-?ria.~roblemsdue to unload in^ of bearings, excessive rates bi weardown, shaft whirling, and gearing misalignment were not rare. Problems were frequently related to the system having too many bearings. In order to better understand the optimum locations for bearings, designers began analyzing shafting as a continuous beam. However. the time reauired to analyze one shafting system confibration by hand calculations precluded complete analyses. The development and general dissemination of digital computer programs, such as reference [23], made it feasible to routinely conduct in-depth studies to optimize shafting systems as well as diagnose recurring problem areas. Factors to be considered in determining the number and location of shaft bearings - are: 1 2 3 4 5

Ship's fixed structure and arrangement. Equality of line shaft bearing reactions. Bearing unit loads and LID ratios. Shafting flexibility./ Lateral vibration natural frequencies (shaft whirl).

Ship's fixed structure such as bulkheads and stanchions will usually require compromises in the shafting arrangement. Also maintenance and overhaul must be considered before final bearing locations are set. From a cost and interchangeability standpoint, all line shaft bearines should be identical. Therefore. the bearings should bve spaced such that the bearing reactions are approximately equal. If this is done, the total number of bearings in the run of shafting is set by the total shaft weight, permissible design unit load, and the acceptable LID limits. The number of line shaft bearings required to support a run of shafting can be tentatively determined a s follows:

where

W

total weight of shafting to be supported (note that gear and stern tube bearings may carry some line shaft weight) p = design bearing pressure (maximum permissible pressure less 5 to 10 psi to allow for variations) based on projected area D = shaft diameter in way of journal (normal practice is to increase the shaft diameter %-% in. in way of bearings) =

PROPELLERS, SHAFTING, AND SHAFTING SYSTEM VIBRATION ANALYSIS '

MARINE ENGINEERING Table 6

Bearing Bearing No.

Fwd SlowS eed Bear 1

Table 7 Tabulation of Bearing Reactions for the Shafting Arrangement Shown in Fig. 1

Bearing Reaction Influence Numbers for Shafting Arrangement Shown in Fig. 1 (Pounds per Mil of Vertical Displacement) After SlowS eed Bear 2

Fwd Line Shaft 3

Line Shaft 4

Line shaft 5

Line Shaft 6

Aft Line Shaft 7

Stern Tube 8

375

Strut 9

Bearing No. Reaction Line-in-Line Cold (as aligned) Hot (as aligned) Weardown Condition5 Measured (Hot) ~

Fwd SlowS eed 8ear

Aft SlowS eed Bear

Fwd Line Shaft

Line Shaft

Line Shaft

Line Shaft

26900 41700 36300 36300

49500 29700 37300 37100 38500

23200 29600 26500 26700 26700

25500 23700 24900 24300 26400

25200 25700 25400 27200 26300

27200 27000 27100 20100 30300

1

2

3

4

5

Line Shaft '7

6

Stem Tube

Strut

8

9

~

...

.

15900 15900 15900 30800 11200

.

62600 62600 62600 51000

...

.

89900 89900 89900 '92200

'

...

a Assume a strut bearing weardown of 0.200 in. and a stern tube bearing weardown of 0.185 in. (consider weardown proportional to bearing pressure).

Notes: The numbers tabulated above represent the effect of raising a iven bearing one mil; e.g., if the forward line shaft bearing is raised one mil, the forward slow-speed gear bearing reaction increaaea 336 lf~,the after slow-speed bearing reaction decreases 493 lb,the forward line ahaft bearing increases 259 Ib, etc.

LID = bearing length/diameter ratio After tentatively selecting the number of bearings and 'pacing them a of the bearing loads under all normal operating conditions is made. The conditions of primary tance are the cold starbup condition, the hot operating condition, and bearing weardown and misalignment conditions. Involved in this analysis is the influence of the shafting on the reduction gear bearing loads or diesel engine bearing loads resulting from the themla1 change the in the position of these when going cold to the hot operating condition. Criteria for aligning the propulsion unit to the shafting are developed on the basis of this analysis. A parametric study of minimum line shaft bearing 'pacing was carried and in reference [241' The conclusion n3ached was that for shafting arrangements having one or more line shaft bearings the minimum span ratio (i.e., ratio of bearing center distance to shaft diameter) should be 14 for shafts with diameters in the range of 10 to 16 in. and 12 for shaft diameters of l6 30 in' The 'pan ratio be in the range of 20 to 22 but the final determination must be and vibraat the based On strength, shaft tion characteristics. 4*4 shafting Calcu'ation Output. The important output from most shafting calculations includes the following in addition to the required shaft diameters:

(a) Line-in-line reactions. (b) Slope of shafting a t discrete poids. (c) Deflection of shafting a t discrete points. (d) Moments in shafting a t discrete points. (e) Lateral natural frequency of shafting. (f) Bearing reaction influence numbers. The significance of the shafting line-in-line reactions (bearing reactions with all bearings set concentric), shaft slopes in way of bearings, shaft deflections, shaft moments, and lateral natural frequency of the shafting is

apparent; however, the importance of bearing reaction influence numbers is not as readily appreciated. Table 6 ;s a tabulation of the bearing reaction influence numbers for the shafting arrangementshown in Fig. 1. The numbers given ih Table represent the change in the magnitude of the bearing reaction of the various bearings as a result of raising any bearing one mil. Thmugh the application of these influence numbers, which reflect the shafting system flexibility, it is possible to investigate the influenceof shafting misalignment caused by thermal expansion, weardown, and &her such effects.Alignmentrequirements are developed on the basis of the Also, the principles bearing reaction inRuence employed with the hydraulic jack method of checking shaft alignment ,(see Section 4.12) originate with the numbers. bearing reaction 4.5 Gear-to-Shafi Alignment. Particular care must be taken in selecting the aftermost and forwardmost line shaft bearing locations to ensure that adequate shafting flexibility is provided. These bearings are subjected to a varying alignment in service. Weardown of the stern tube bearing alters the load on the aftermost bearing; the thermal expansion of the reduction gear structure changes the load on the foMiardmost bearing. These * effectsmay be noted from the tabulatioh given in Table 7. when the propulsion plant goes from the cold to the operating condition, the slow-speed gear besings may to 30 mils relative to the line shaft bearings. rise his rise can significantly alter the reactions of the slowspeed gear bearings and the forward line shaft bearings. Of particular concern is the fact that the static load on the forward slow-speed gear bearing decreases while that on the after bearing increases,. As can be seen from the typical reduction gear bearing reaction diagram shown in Chapter 9, this causes the slow-speed gear to assume a crossed-axis position relative to the slow-speed pinions which are not similarly affected. As a result, the tooth load will tend to be more heavy on one end of each helix. Reference [25] contains a detailed discussion of the effects

of unequal gear bearing reactions on gear performance and the maximum permissible differences between the gear bearing static loadings. The gear manufacturer should state the maximum difference permitted between the static loadings on the slow-speed gear bearings (see Chapter 9). The alignment in the athwartship direction should be such that no significant forces are imposed on the slowspeed gear bearings in the horizontal plane. The allowable setting error is conveniently used as an index of shafting flexibility in way of the reduction gear. The allowable setting error, ASE, is defined as the allowable difference in the static vertical gear bearing loads divided by the difference between the bearing reaction influence number of the forward slow-speed gear bearing on itself and the after slow-speed gear bearing on itself. Therefore, the ASE is determined as follows:

*ASE =

AR I11

-

I22

(11)

AR = allowable difference between two slow-speed Ill

I22

gear bearing static reactions reaction influence number of forward slowspeed gear bearing on itself = reaction influence number of aft slow-speed gear bearing on itself =

The ASE number represents the total of the error permissible in estimating the thermal rise of the slowspeed gear bearings relative to the line shaft bearings and the error permissible in setting the gear to the line shafting without exceeding the maximum allowable difference in the static slow-speed gear bearing reactions. An absolute minimum acceptable value for the ASE has been recognized to be f.010 in.; see reference [26] for additional discussion of this parameter. If the flexibility of the shafting meets the ASE criterion, the analysis proceeds to an investigation of the gear-to-shaft alignment. Beginning with the line-in-line reactions, that is, the bearing reactions with all bearings concentric, and with the estimated thermal rise of the gear bearings relative to the line shaft bearings when going from the cold to the operating temperature, alignment data are established which will provide approxi-

mately equal slow-speed gear bearing static reactions when in the operating condition. I t must additionally be ascertained that the line shaft bearing reactions are satisfactory under all operating conditions. Since the actual positions of the slow-speed gear bearings are difficult to ascertain relative to the line of shafting, their positions are determined by measuring the drop and gap of the slow-speed gear shaft flange";elative to the line shaft flange. Drop is the vertical distance between the centers of two adjacent flanges; gap is the differencein opening between the top and bottom of the two flanges (nonparallelism of the flange faces). It is, of course,-necessaryto know the position of the line shaft flanges relative to the line shaft bearings. One alignment technique is to support the line shafting sections so that the centers of the flanges are concentric with the bearing centerline and the flange faces are perpendicular to the bearing centerline; this is accomplished by s u p porting the sections at approximately the 2/9 points from each end. When this is done, the drop and gap are measured directly a t the flanges. The positions of the slow-speed gear bearings $elative to the centerline of the line shaft bearings are readily determined from the drop and gap of the slow-speed gear shaft flange relative to the line shaft flange; this is accomplished by means of simple geometry. With these data, the bearing reaction influence numbers can be used to plot the bearing loads for various alignment conditions. Such a plot is shown in Fig. 9. Figure 9 is an informative means of illustrating the effect of thermal rise of thevslow-speed gear bearings and the effect of alignment errors in addition to bearing loads. Table 7 shows the cold alignment bearing reactions with an alignment corresponding to point A on Fi& 9. The hot reactions listed in Table 7 are the bearing reactions estimated after the gear has reached operating temperature; this is point B on Fig. 9. Finally, bearing reaction influence numbers provide a means to study the effects of bearing movements from hull deflections and bearing weardown. An analysis of bearing reactions with the stern tube and strut bearings worn down is given in Table 7. 4.6 Propeller-to-Shaft Interface. Design details of

I'

I I

I

I

I

Fig. 9

I

Gear-to-shaft alignment analysis

i'

the propeller-to-shaft interface are a critical aspect of a shafting system design. DuringWorld War I1 and earlier years, propeller shaft failures in way of the propeller were not rare; and difficulties have been experienced in more recent times [20, 211. However, the advances in design technology (e.g., stress relief grooves a t the forward end of the propeller and the aft end of the liner, shortened-and spooned keyways, slotted keys, and improved sealing methods) have significantly improved the reliability of propeller shafts and increased their service lives. Also, improvements in inspection technology have provided the means to detect incipient cracks and thus have greatly reduced the loss of propellers at sea. Details of the propeller-to-shaft interface required for naval ships are specified by reference [27]. The naval type of propeller-to-shaft interface is consistent with merchant praltice and is reported to have a comparable service history. The propeller keyway is a stress concentration and a weakening factor to the shaft even though the keyway has generous fillet radii and the forward end of the key is slotted to relieve the key load at the forward end. For this reason, propeller nuts have been developed which incorporate annular pistons moved by hydraulic oil or grease. These "hydraulic" nuts provide the means to apply large forces of known magnitude to the propeller, pushing it onto the shaft taper such that no propeller key is required. Keyless propeller designs rely entirely on the friction between the hub and the shaft to withstand the propeller torsional and thrust loads. Conse-

quently, the contact pressure, material stresses, shaft taper, propeller-shaft interface friction coefficient, and the push-on force must be thoroughly engineered. Although some keyless propeller designs use oil pressure to expand the propeller hub while the propeller is forced onto the shaft taper, most designs call for the mating surfaces to be thoroughly degreased before the propeller is fitted to the shaft taper to ensure that the highest coefficient of friction is obtained. However, it is maintained by some that when oil is used to expand the hub for fit-up, shortly after the pressure is relieved the oil layer is squeezed out of the propeller-to-shaft interface, resulting in a friction coefficient equal to that when the propeller is pushed up dry. Most designs use a hydraulic pressure in the propeller-to-shaft interface a t least equal to the calculated radial pressure to expand the hub for removing the propeller. For a dry, greaseless, installation the coefficient of friction may vary from a low of approximately 0.13 to a high of approximately 0.1s. The value of the effective friction coefficient is dependent on the percentage area of the mating faces which are in contact at initial fit-up (prior to forcing the propeller up the shaft) and the ma-, terials in contact. Some keyless propeller designs incorporate a cast-iron sleeve in the propeller hub bore because tests have shown that higher coefficients of friction can be obtained this u-ay. Using an intermediate nodular iron sleeve with an initial surface contact area between the sleeve and shaft taper of 95 percent or more, a coefficient of friction of 0.18 can be achieved. When the

t

1

I

1

I

I I-

initial surface contact area is approximately 50 percent, a coefficient of friction of 0.13 can be expected. When a bronze alloy propeller is fitted to a steel shaft without using an intermediate cast iron sleeve, the maximum coefficient of friction obtainable at the propeller-shaft interface is a~~roximatelv 0.15. The shaft taper used with keyless propellers generally falls within the range of 1:12 to 1:20. The shaft taper should be selected such that the sum of the maximum astern thrust and the axial push-off force due to the taper does not exceed the axial frictional resistance. This consideration will limit the maximum shaft taper. On the other hand, if the shaft taper is too small, the ahead thrust may force the propeller up the taper and overstress the hub. 4.7 Cold-Rolling. The development of fretting fatigue cracks in propeller shafts a t the forward end of the propeller hub and a t the after end of the shaft sleeve is one of the most common modes of propeller shaft failures. Although surface cold-rolling will not eliminate the occurrence of fatigue cracks, cold-rolling of propeller shafts for a distance forward and aft of the forward end of the propeller shaft taper and in way of the ends of the liners has been shown to be an effective means of retarding the propagation of fatigue cracks [28]. 4.8 Protection from Seawater. Except in the case of designs in which all bearings are of the oil-lubricated type, outboard shafting involves the use of sleeves which are shrunk on the shafting in way of bearings, stuffing boxes, and fairings. Shaft sleeves are made of bronze or other materials which are resistant to attack by seawater. Ships having a single short section of outboard shafting employ a single continuous sleeve. Where continuous sleeves are not used, the sections of shafting not covered -by sleeves are protected by applying a rubber [29] or plastic [30] compound directly to the shafting surface. The adequacy of both rubber and plastic protective coverings for outboard shafting has not been unifornily good. The protection offered to outboard couplings by such coverings has been particularly unsatisfactory on occasions. Rotating coupling covers (fairwaters), which clamp onto and rotate with the shaft, thereby eliminating the violent erosive flow of water around coupling bolts, have been used to avoid the deleterious effect of the water. A reliable static sealing arrangement a t the propeller, which prevents seawater from contacting the propeller shaft, is of the utmost importance. A propeller-shaft assembly which allows seawater to contact the shaft will not have an endurance limit and therefore it is only a matter of cycle accumulation before a failure occurs. Details concerning propeller-hub sealing arrangements, which are necessary with systems utilizing water-lubricated stern tube bearings, are contained in reference [31]. 4.9 Shaft Couplings. Except in instances where special considerations preclude their use, shafting sections are connected by means of integrally forged couplings as illustrated on the line shaft section shown in

,,,

TYPICAL BOLT

SHAF

FT

LSPLIT COLLAR Fig. 10

Removable flange coupling

Fig. 10. Although the design of virtually all integral shaft couplings is similar, details of shaft coupling designs can vary considerably. For example, despite individual preferences, no specific number of coupling bolts has been established as optimum and the proportions of flange dimensions may vary from one design to the next. Guidelines for the design of flange couplings for merchant vessels are given in classification society rules such as reference [18] and similar guidelines for naval vessels are given in references [22,27]. An effort to standardize shafting couplings was made in reference [19], and SNAME T&R Bulletin 3-20 (Guide for the Design of Line Shaft Couplings) was subsequently prepared. Couplings with removable flanges are required in some instances; for example, those cases where a liner must be installed on a shaft which requires a bolted flange on each end. Figure 10 illustrates a typical removableflange coupling and shows the means provided to transmit both thrust and torque. Both torque and thrust are normally transmitted by friction between the shrunk-on muff and the shaft. The keys are a backup for the transmission of torque and the split collar is a backup for the transmission of thrust. On occasions shafting arrangements are designed such that it is necessary to remove the forward flange of a stern tube shaft in order to withdraw the stern tube shaft aft; this is not a preferred arrangement because it is most difficult to remove a flange without damaging the flangeshaft interface. A preferred arrangement is one in which the stern tube bushings are made sufficiently large so that they can be removed to permit the stern tube shaft to be withdrawn aft with the forward flange in place. 4.10 Shaft Axial Movements. Axial movement of the shafting relative to structure that is fixed to the hull must be considered to establish proper clearances between the propeller and stern frame structure and the clearances between bearing housings and rotating elements secured to the shaft within bearings (oil slingers, oil disks, etc.) There are four factors which contribute to the movement of the shaft relative to hull structure; these are:

378

MARINE ENGINEERING

PROPELLERS, SHAFTING, AND SHAFTING SYSTEM VIBRATION ANAtYSlS

effect a static balance. Good practice dictates that adjoining shafting sections be installed such that the residual static unbalance, as determined by a check on the rails after final machining, tends to cancel. Although shafting sections have occasionally been specified to be dynamically balanced (shaft sectiorls SHOWN IN FIG. I rotated in a balancing machine to determine both static and dynamic unbalance), there are conflicting schools of thought regarding the necessity of a dynamic balance. FIRST MEASUREMENT It has been argued that the tolerances customarily im+ SECOND MEASUREMENT posed on the manufacture of shafting sections in conjunction with good shop practice precludes objectionable shafting unbalance. 4.12 Determination of Shaft Alignment. There are basically two ways that the alignment of an installed and complete shafting system can be checked. One, which is akin to the drop-and-gap method of alignment a t initial'installation, is to remove the bolts from a coupling and compare the relative position of the two o 1000 2000 3000 4000 SOW flanges with the calculated value. The second method, JACK PRESSURE, PSIG which is both easier to accom~lishand more meaninnful. , is the so-called hydraulic jack method. F i 11 baring reaction determined by hydraulic lack With the hydraulic jack method of checking shhfting alignment, a calibrated hydraulic jack is used-to determine the actual load Y U D D O bv ~ ~ a~ bearinn ~ and this actual load is then compared with the desired gad. The 1 Thrust bearing clearances. Axial clearances beload is determined by placing a hydraulic actual bearing tween the thrust ccrlla~and shoes permit a corresponding jack as close to the bearing housing as possible (bearing fore-and-aft movement of the entire shafting system. foundations are often designed with an extension to 2 Axial deffections. The propeller thrust results in a provide a jack foundation). A dial indicator is located small axial deflection of the shafting and thrust bearing. immediately above the jack so as to measure vertical 3 Temperature difierences. The shafting can be a t movement of the shaft. Where possible, the anchor a wa& temperature (70-80 1;") relative t o that of the point for the dial indicator should be independent of hull structure (about 30 F). the bearing housing. Before recording any readings, the 4 Hogging and sagging induces bending strains in shaft should be lifted at least once to ensure that the the hull wWe the shafting is not similarly strained. shaft can be lifted 20 to 30 mils without coming into This is conveniently amessed by assuming an extreme contact with the upper half of the bearing; this prefiber hull bending strem and the neutral axis of the hull liminary jacking tends to reduce hysteresis in the shaft in bending; the 8tress, and corresponding hull strain, a t and erratic readings. For short shaft spans, a dial the s h d t centerline ia then determined by interpolation. indicator should also be installed on adiacent bearings The foregoing factors would generally not reach maxi- so that any rise of the shaft in these dearings can be mum values simultaneously, but they are prudently noted. At a later time, this may help t o explain unconsidered to do so. Typical axid movements of the predicted readings. With the dial indicators and jack in place, the shaft is propeller (the point a t which movement is a maximum) d a t i v e t o the hull are 0.5 in., for tankers with very short raised and lowered in increments, noting the jack load shafts, to 2 in., for ships with long shafts. corresponding to each increment of shaft rise. These 4.1 1 Shafting Balance. Solid shafting is inher- data are plotted as shown in Fig. 11. The data points ently bdmced, but hollow shafting requires attention in will conform to two basic slopes. The slope of the liftthb regard. The bdanee of hollow shafting is accom- versus-pressure curve as the load is transferred from the plished during the machining operation by shifting lathe bearing to the jack represents the spring constant of eentem prior to the fina.1 machining cuts. The amount the bearing shell, bearing housing, and the like. Wllerl the of unbalance in a shaft can be determined by either a shaft lifts clear of the bearing, an abrupt change in static OI dynamic balancing technique. the slope of the data points occurs. The second slope After the rough machining cuts have been made, a corresponds to the bearing reaction influence number shaftipg section can be ataticafly balanced by removing for the bearing. Due to friction in the shafting and jack system, the the shaft seetion from the lathe, placing it on rails, noting the equilibrium position of the shaft section, shifting the data points when raising and lowering the shaft do not lathe centers to compensate for the unbalance, and then coincide, the result being the equivalent of a hysteresis taking additional machining cuts on the shaft section t o loop. The deflection-versus-load plot will show a lower

shaft lift at a given jack load for the increasing load curve that the bearing should be raised or lowered should be than for the decreasing load curve. Experience indi- based on the calculated influence numbers rather than cates that the true relationship between the jack load the influence numbers determined by jacking. When jacking bearings that are very close together and and shaft lift is approximately midway between the lines determined when raising and lowering as indicated by in cases where the jack must be located some distance Fig. 11. However, in cases where the increasing and from the bearing, the jack load should be multiplied by decreasing load lines are significantly different, the mean a correction factor to obtain the load a t the bearing. The correction factor is as follows: line should favor the increasing load line. With the mean line representing the true relationship between the jack load and shaft lift established, the load which would be on the jack a t zero shaft lift and with the bearing removed is determin$ by extrapolating the mean line downward to zero shaft lift. Since the jack where and bearing are close together, the load as determined I& = influence of bearing on bearing can also be considered as the load on the bearing if the l i b = influence of jack on bearing jack were removed (or the bearing load being sought). Under favorable jacking conditions (no binding of the These influence numbers are determined by including shaft in the bearing due to athwartship misalignment, both the jack and the bearing being jacked as support interference with stuffing boxes, etc.) experience shows points in the shafting system calculations. To be that the accuracy of the bearing reactions determined is theoretically accurate, this correction factor should be usually within 10 percent. However, the influence num- used for every bearing that is jacked; however, only in bers obtained by jacking may not be as accurate. When the aforementioned two caqes is it a factor of significance. the bearings being jacked are located towards the middle Table 7 contains a tabulation of the measured bearing of the shaft and span lengths are fairly equal, jack influ- reactions for the shafting system in Fig. 1 and illustrates ence numbers are generally within 30 percent of the typical jacking results. The oil in the reduction gear calculated influence numbers. For bearings located near was heated and circulated a t operating temperature; the ends of the shaft, the influence numbers obtained by therefore, the measured reactions should be correlated jacking may disagree with the calculated values by with the hot reactions. 50 percent or more. The hydraulic jack procedure can also be used to Both the load and influence number errors aye due to detect bent shafts in that the bearing reactions can be inaccuracies which are inherent in the jacking procedure; determined with the shaft rotated in 90-deg increments. e.g., the jack not being located at the bearing center, the If the bearing reaction changes significantly with shaft load center in adjacent bearings shifting as the shaft is position, a bent shaft can be suspected. This technique raised, and hysteresis in the shafting system. Conse- is very useful when analyzing a shaft that is suspected quently, when a bearing is to be realigned, the distance of being bent.

CALCULATED

MEASURED

-

379

Sectiwn 5 Bearings 5.1 Introduction. Main propulsion shafting is supported by bearings which maintain the shafting in proper alignment. These propulsion shaft bearings divide themselves naturally into two groups; those bearings inside the watertight boundary of the hull and those bearings which are outside the hull watertight boundary. The requirements imposed upon the design of main shaft bearings are extremely severe. The bearings are required to operate a t speeds ranging from 0.1 rpm, when on jacking gear, to 100 or more rpm in either direction of rotation. And, unlike some applications, the bearing loads do not vary with rpm but are essentially constant at all speeds. Reliability is heavily emphasized'in the design of bearings because there is no redundancy for bearings and a single bearing failure may incapacitate the propulsion system. I n addition to the radial bearings which support the

main shafting, there is located inside the ship a main thrust bearing which transmits the propeller thrust from the shafting to the hull structure. Figures 1 and 2 show the two typical main thrust bearing locations. Often, the main thrust bearing is designed as an integral part of the main engine and is provided by the main engine manufacturer. l'or details concerning main thrust bearings, see Chapters 9 and 20. 5.2 Line Shaft Bearings. Bearings located inside the ship's water-tight boundary are called line shaft bearings, although they are sometimes referred to as steady or spring bearings. Almost without exception, these bearings are ruggedly constructed, conservatively designed, babbitt lined, and oil lubricated. Except in special cases, the bearings are self-lubricated by rings or disks arranged in such a manner that lubrication is effected by the rotation of the shaft. Roller bearings

MARINE ENGINEERING OIL SCRAPER

7

rANTI- ROTATION DOWEL

PROPELLERS, SHAFTING, AND SHAFTING SYSTEM VIBRATION ANALYSIS r

WINGED INSPECTION COVER

HALF

\

,-OIL

LUBRICATING RING 7

COOLING COIL CONNECTION

Rg. 12 Self-aligning line shaft bearing with oil disk lubrication

have been used in the smaller shaft sizes, but the advantages of lighter weight and lower friction have in general not been sufficient to offset the higher reliability and lower maintenance costs of the babbitt-lined type. Line shaft bearing housings are made of steel castings or fabricated of steel plates welded together. Completely satisfactory bearing housings are obtained by either method, and manufacturing costs govern the construction method used. Since rigidity is of more concern than strength, low carbon steel is used as the material for bearing housings with the exception of bearings for naval combatant vessels, in which case high-impact shock requirements may necessitate the use of high-strength steel. Bearing housings are split horizontally a t the shaft centerline. The bottom half of the bearing must be very ruggedly designed since it carries the vertical shaft load and any side load that exists. The bearing housing supports a h h v y steel removable shell which is limed with babbitt. The shaft rests on the babbitted surface. The bearing shell can be made with a self-aligning feature by providing a spherical or crowned seat at the interface between the bearing shell and housing. This allows the axis of the bearing shell to align exactly with that of the shaft. Figure 12 is a section through a bearing with a self-aligning feature and Fig. 13 is a section through a bearing that is similar but without a self-aligning capability. The general construction of bearing housings and shells can be observed from Figs. 12 and 13. Except for the aftermost line shaft bearings in mer-

chant applications, it is general practice to babbitt only the bottom half of the bearings since these bearings would never be expected to be loaded in the top. However, the aftermost bearing (the one closest to the stern tube) may become loaded in the top particularly when the stern tube bearing is water-lubricated. Waterlubricated bearings are subject to a large amount of wear which can result in severe misalignment. It is considered good practice to provide the maximum practicable amount of babbitt in the top half of the aftermost line shaft bearings when water-lubricated stern tube bearings are used. With oil-lubricated stern tube bearings, the probability of the after bearing becoming loaded in the top is considerably reduced. . Babbitt that is centrifugally cast onto the bearing shell is considered preferable to that which is statically poured. The former technique dependably provides a more secure bond between the babbitt and the bearing shell. The desire for centrifugally cast babbitt has required some adjustments in bearing shell and housifig design. Babbitt can be of either the lead or tin base type. Tin-base babbitt has greater strength and is generally preferred for shaft bearings; it is specified almost exclusively for centrifugally cast bearings. Lead-base babbitt is preferred where embedding, conforming, and antifriction are primary considerations. Lead-base babbitt has a lower yield point and a slightly better fatigue resistance. Physically the load-carrying length of the bearing

Fig. 13 Nan-self-aligning line shaft bearing with oil ring lubrication

i

1

should not exceed two times the shaft diameter (LID = 2) nor should it be much less than one shaft diameter (Ll D = 1). These limitations are set to assure uniform bearing contact in the case of long bearings and to prevent excessive end leakage of the oil from impairing adequate lubrication for short bearings. The oil reserv.oir must be sized to operate during extreme roll and pitch conditions without leaking oil by the shaft or disabling the bearing lubrication system. Furthermore, the oil quantity and sump surface must be sufficient to dissipate the heat generated. Line shaft bearings are sometimes designed with cooling coils located in the sumps as shown in Fig. 13; however, experience has shown that the cooling coils are rarely, if ever, needed. Line shaft bearings may be lubricated by means of oil rings, an oil disk, or by a supply of oil under pressure (wick-lubricated bearings have fallen into disuse). Ring oil-lubricated bearings contain two or three metal rings with a diameter of 1.25 to 1.5 times that of the shaft (the ratio decreases with larger shaft diameters). The number of rings in a bearing should be selected such that no ring is required to distribute oil for an axial distance g,reater than 7 in. on either side of the ring. The rings rest on top of the shaft and dip into an oil reservoir located beneath the bearing shell. Figure 13 is an example of a ring-lubricated bearing. As the shaft turns, the rings are rotated by the frictional contact with the top of the shaft. Oil which adheres to the ring in way of the oil reservoir is then carried up to the top of the shaft

where, a part of the oil is transferred to the shaft and subsequently carried into the contact region of the bearing. Ring-lubricated bearings have proved to be capable of accommodating large angles of list and trim and have proved to be reliable in service with design bearing unit loads of 45 psi. With regard to the possible adverse effects of trim, tests have been conducted which demonstrated that ring-lubricated bearings can accommodate angles of approximately 10 deg from the horizontal with no sacrifice in performance. Reference [32] discusses the performance of oil rings based on laboratory tests and shows, among other things, the sensitivity of the quantity of oil delivered to the oil viscosity. Disk-lubricated bearings use a metal disk clamped to the shaft a t one end of the bearing shell. The disk may have a flange as illustrated by Fig. 12. As the shaft turns, thelower portion of the disk, which is immersed in , coated with oil. This oil is carried to an oil r e s e ~ o i r is the top where a metal bar scrapes the oil from the disk and guides it into passages where it is admitted to the top of the shaft and then into the contact region of the bearing. Disk-lubricated bearings have been successfully applied with design unit pressures of 75 psi. I n special cases, line shaft bearings may be lubricated by oil supplied by a pump. If the shafting system is very long, sump pumps are required to return the oil from the bearings since a gravity drain is not feasible considering the possible trim and pitch conditions of the ship. While this method of lubrication assures an adequate supply of oil a t all shaft speeds, it has the dis-

382

PROPELLERS, SHARING, AND SHAFitNG SYSCEM VIBRATION ANALYSIS

MARINE ENGINEERING

advantage of the extra pumps and complexity. Furthermore in the event of pump failure, the bearings may be damaged from the lack of oil. The load which can be supported by a babbitted journal bearing is dependent upon the method of lubrication, the bearing length to shaft diameter (LID) ratio, and of course the installation workmanship. I n the past, babbitted journal bearings were restricted to bearing pressures of 20 to 30 psi based on the projected bearing area. This limitation on allowable pressure resulted in bearings with LID ratios as large as 2. Even with high LID ratios, the shafting systems had very closely spaced bearings such that the bearing loads were very sensitive to alignment. The use of higher bearing pressures along with the use of more sophisticated techniques in positioning bearings has resulted in more reliable shafting systems by virtue of the more favorable bearing LID ratios and more flexible shafting systems. The most severe demands on the lubricating system of a line shaft bearing do not correspond to full-power, fullrpm operation, but to the condition when the shafting is rotated by the turning gear a t about 0.1 rpm for extended periods of time to facilitate uniform cooling or heating-of the main turbine rotors. If the lubrication system fails to deliver adequate oil to the journal under this condition, the oil film which separates the bearing journal from the babbitt will not be replaced as it is squeezed out; consequently, metal-to-metal contact and damage to the bearing surface may occur. Lubrication provisions have a strong influence on a bearing's ability to operate satisfactorily in the critical jacking mode of operation; and, consequently, the means of lubrication .strongly influences the extent to which line shaft bearings can be loaded. As a guide, it has been found that as little as 25 drops of oil per minute on the journal surface is adequate to sustain indefinite operation in the jacking mode a t bearing pressures of about 75 psi. With proper attention given to design details, ringlubricated bearings, disk-lubricated bearings, and pressure-lubricated bearings can carry increasingly higher unit loads in that order. Disk-lubricated bearings can carry a higher unit load than ring-lubricated bearings based on the assumption that the oil scraper functions properly. Very close controls must be maintained in the manufacture of oil scrapers because manufacturing flaws which are hardly perceptible can have a large influence on their performance. 5.3 Outboard Bearings. Outboard bearings can be further classified as stern tube or strut bearings. Figures 1 and 2 show the locations of these bearings relative to the ship arrangement. Outboard bearings can either be water lubricated or oil lubricated. Almost without exception in this country, all outboard bearings were water lubricated up till about 1960 when a transition to oil-lubricated bearings began. This transition to oil-lubricated bearings was stimulated by the unduly short service life of many of the water-lubricated bearing assemblies during that period. It is believed that the shortened life of the

water-lubricated bearings was caused by the larger ship sizes which had greater bearing loads and more contaminated water passing through the bearings (larger ships operate with less clearance between the hull and channel bottoms such that more silt, mud, and sand is drawn into the bearing surfaces). Reference [33] reports the experience of ship operators regarding stern tube bearing wear. Minimization of vibration was also influential in the promotion of oil-lubricated bearings. Particularly with larger and fuller ships, variations in the water inflow velocity to the propeller generate large variable bending forces on the shafting; many instances of pounding of the shafting in the forward stern tube bearing and the stern tube stuffing box of single-screw ships have been noted particularly when five-bladed propellers were being used. Oil-lubricated bearings which have close bearing clearances eliminate-the pounding and associated maintenance of propeller shafts and stuffing boxes. Oil-lubricated stern tube bearings also reduce the power losses in the shafting system. For a 22,000-shp ship an efficiency improvement of about 1.5 percent can be expected with oil-lubricated vice water-lubricated outboard bearings. Although oil-lubricated outboard bearings are favored by many, water-lubricated bearings remain in common use. Figure 14 illustrates a typical water-lubricated strut bearing design. A water-lubricated stern tube bearing design is similar ;xcept 'that the bearing bushing is fitted inside the stern tube rather than the strut barrel. Water-lubricated bearings basically consist of a bronze bearing bushing which retains a number of bearing contact elements that may be made of either lignum vitae wood, phenolic composition, or rubber bonded to brass backing strips. A sleeve is installed on the shaft to provide a corrosion-resistant contact surface. Careful consideration must be given to the selection of the liner material in relationship to the bearing material [34]. When brass-backed rubber strip bearings (rubber stave bearings) are used, as is common in naval practice, dovetailed slots are accurately cut in the bushing to accommodate the bearing staves. Sufficient metal is left between each slot to hold the staves securely; the space between staves also provides a cooling water flow ~ a s s d e . As indicated by Fig. 14, bearings employing lignum vitae and phenolic materials are similar to rubber stave bearings. A "V" or "U" shaped groove is cut a t the longitudinal joints of the blocks to provide lubricating and cooling water flow. Brass retaining strips are generally placed at four points around the circumference to secure the contact elements. Lignum vitae and phenolic materials absorb water and consequently tend to swell. Phenolic materials are usually installed when dry and consequently swell significantly when put into service. Swelling must be considered in the design of both lignum vitae and phenolic bearings. Lignum vitae must be kept damp a t all times as it will otherwise become dry and crack. Reference [35]

FhClLtTATE SMAFT WlTHDRAWI)L 1--

\

ATHWARTSHIP V t E I TAKEN AT 'A-A'

F~~RWATER

4

\v

ALTERNATE DESRN FOR LL6UUM VIThL OR PHLWOLtC eOMPOSITIQLIBLOCUS

383

f/

FORE a AFT VIEW TAKEN AT "B-B* RUBBER

i

PEAK BULKHEAD

SEAL SLEEVE

Fie. t 5 TweuI

aif-hbticcrled rtacn hrbe bearing

conhim details pertinent to the application of lignum vitae ta prapelIer shaft bearings. Water-lubricated bearings are designed with L / D ratios of the order of 4 for the bearing adjacent t o the prop&er and 2 for those forward of the prnpdh3r bearing. GIassifieation societies often specify these values as minimum lengths. Unit loadings of the bearings based on prcrf&ed area (shaft diameter times hearing length) are normally under 40 psi; however, great care must be taken in placing importance on the absolute value of bearing eontact pressures which are based on the projected area. Not only does the eccentricity of propeller thrust alter the loading but also the load distribution is both difficult to assess and is subject to radical .change.

Outbwrd baring materials msy w a r 0.2 to 0.5 in. M o r e being replaced. O u t b o d hearings are occasionally aligned to ZL slope corre8pnding ta the static slope of the shaft in way of the bearing in order to obtain mom uniform bearing eantmt when initially placed in service. However, this procedure has not proven entirely satisfactory from a we&down standpoint since only the starting point of the wear process is changed and there may be little influence on the ultimate wear pattern. OiMuhricated beprings, as illustrated by Fig. 15, have been used in stern tubes and bossings more so khan struts. This Is partly because of the fewer bearing problems with water-lubricated strut bearings and partly due t o the

3 84

PROPELLERS, SHAFTING, AND SHAFTING SYSTEM VIBRATION ANALYSIS

MARINE ENGINEERING

),

T FILL

Fv7

CONNECTION ~

~

~

I 0 FT. A8OVE LOAD WATER LlNE LOW LEVEL ALARM

GRAVITY LlNE

THERMOMETER CONNECTION -DRAIN CONNECTION

Fig. 16 Stern tube lubricating oil diagram

difficulties in adapting the system to struts. Oil-lubricated bearings do not require a liner to be installed on the shaft since contact with seawater does not occur nor is there any significant shaft wear. Also, no bushing is inserted in the stern tube; the bearing shells, which have heavy wall thicknesses, are pressed directly into the stern tube. The L I D ratios of the heavily loaded after

stern tube bearing have ranged widely. Early designs ~ ~ ~ ~ ~ ~ had ratios of 2.5 but a trend toward a value of 1 was subsequently established. Oil-lubricated stern tube bearings are totally submerged in oil, and seals on the after and forward ends of the tube prevent the ingress of seaw3ter and the leakage of oil into the ship, respectively. iThe pressure of the oil in the stern tube is maintained above that of the ambient seawater by means of a head tank which is located about 10 f t above the full-load waterline. Ships which have large draft changes may require two head tanks; one for full-draft operation and one for ballast operation. Figure 16 illustrates a typical lube oil diagram for an oil-lubricated stern tube bearing. A small pump is usually installed M shown to force oil circulation through the stern tube. The oil flow is such that oil is circulated through both bearings. Many variations of this system have been used including the deletion of the pump; owners often specify filters, heaters, coolers, and coalescers to condition the oil as it passes through the circuit. Coolers are rarely used as the temperature leaving most stern tubes does not exceed 120 I?. Although the unit bearing pressure based on the projected area normally falls in the 70-psi range for oillubricated bearings, the actual operating pressure is probably closer to twice this value. An inspection of the bearing contact area after operation reveals that the after bearing is loaded only on the after end for a length of about one shaft diameter; shorter bearings are often advocated for this reason.

Section 6 Propellers 6.1 Introduction. Very early in the preliminary design spiral, the hull resistance and propulsion system must be established. Section 4 of Chapter 1 deals with the considerations involved in developing the main propulsion system. The propulsor, a device which converts engine torque to ship thrust, can be one of the important determinants of the type of propulsion plant employed. Because of the interfaces between the machinery, hull, and propulsor, the design of the propulsor is usually a task undertaken jointly by a naval architect and a marine engineer. The responsibility for the design of the propulsor varies from one organization to the next, but one approach is to assign the naval architect the responsibility of developing the hull lines and the propulsor hydrodynamic design; he is supported by the marine engineer who provides the proper interface between hull and machinery to assure that an optimum overall propulsion system is obtained. This is the general basis upon which the Society's two publications, Principles of Naval Architecture and Marine Engineering, have been written. Accordingly, details regarding the propulsor hydrodynamic design are not covered in this text.

Figure 4 of Chapter 1 compares the optimum efficiency values for a number of different types of propulsors. This information gives guihnce as to the relative merits of one propulsor versus another from an efficiency standpoint. However in the preliminary design stage, more specific information is required in order to make the necessary trade-off studies to support a design selection. Systematic model tests of propulsors provide the necessary information for the trade-off studies, and in many cases the final design. Reference [I] provides the information necessary to carry out in-depth propulsor studies. Section 6.2 provides a description of the mechanical aspects and performance characteristics of the various types of propulsors to aid in defining the circumstances involved with the possible alternatives. Most ship propulsors are of the propeller type; therefore, Section 6.3 provides detailed guidance regarding the preliminary selection of the characteristics of ship propellers. Propeller manufacturing tolerances, which a marine engineer is frequently called upon to control or specify, are discussed in Section 6.4.

~

,

6.2 Propulsor Types. As noted in Chapter 1, the L y type of propulsor to be used must be selected very early in the ship design process as the type of propulsor can have a strong impact on the design of the ship itself. The vast majority of ship propulsors are of the solid fixed-pitch propeller type. Nevertheless, there are a number of other types of propellers which may be more suitable in particular instances. A brief description of the mechanical aspects of the various types of propellers is as follows: Canventional. Fixed-Pitch Propellers. Most propellers are of the 'Conventional fixed-pitch type and are made from a single casting. Conventional fixed-pitch propellers usually have an efficiency, cost, and simplicity advantage over other types of propellers. Detachable-Blade Propellers. Detachable-blade (or built-up) propellers consist of a separately cast hub and blades. The blades are bolted to the hub to form the composite propeller. When operating conditions are such that there is a great probability of propeller blade damage, detachable-blade propellers offer the advantage that individual blades can be replaced. Also some blade attachment designs have elongated bolt holes which offer the advantage that small modifications in pitch can be made, which permits adjustment in the operating rpm. The disadvantages associated with detachable-blade propellers, as compared with propellers made from a single casting, are the greater first cost, greater complexity, and inherently lower propeller efficiency (resulting from the larger hub). Controllable- and Reversz3le-Pitch Propellers. The blade angle (propeller pitch) on a controllable- and reversiblepitch (CRP) propeller can be controlled remotely. This type of propeller is advantageous in any of the followinp: situations:

1 Where the operating conditions vary widely (such tug and trawler applications). 2 Where diesel engines or gas turbines are used for propulsion (the CRP propeller permits adjustment of the engine rpm-power relationship to provide gr~ateroperating flexibility). 3 Where reversing capabilities are not readily obtainable from the main engine (e.g., gas turbines). 4 Where rapid or frequent changes in the direction of thrust is a desirable capability. M

At the propeller design point, the efficiency of CRP propellers approaches the efficiency of fixed-pitch propellers. The larger hub of a CRP propeller prevents its efficiency from exceeding that of a fixed-pitch propeller. Off the design point, the CRF propeller efficiency is less than that of a fixed-pitch propeller designed for that operating condition. This is because all sections of a CRP propeller blade are rotated through the same angle as the pitch is changed; thus, the angles of attack of the various blade sections along the propeller radii are optimum only a t the design point. Pitch changes are controlled remotely. The torque required to turn the blades and hold them in position is

385

obtained by controlling the oil pressure on a piston which is mechanically linked to the propeller blade throws. The hydraulic pistons are located in the propeller hub or in a section of shafting located inside the ship. For a description of a typical CRP installation, see reference [36]. Propellers in Nozzles. There are two types of arrangements which fall into this category; namely, the pump jet and the Kort nozzle. In the pump jet arrangement the propeller is placed in a rather long nozzle with guide vanes either forward, aft, or both places relative to the propeller. The pump jet is normally considered where propeller noise is important. Due to the resistance of the nozzle and guide vanes, the overall efficiency of the pump jet arrangement is strongly dependent on particular circumstances. Kort nozzle propeller arrangements show efficiency advantages in applications where the thrust loading is high; examples of such applications are tugs, trawlers, and large slow-speed ships (see Fig. 4 of Chapter 1). The Kort nozzle arrangement consists of a propeller located in a nozzle of relatively short length (the length/ diameter ratio of the nozzle is in the range of 0.5 to 0.8). Kart nozzles are extensively used in connection with tugboats because the bollard pull and towing pull can be increased 30 to 40 percent as compared with a propeller operating alone without a nozzle. Tandem Propellers. As the horsepower requirements for a ship increase, a single propeller can become inadequate due to restrictions on the propeller diameter, draft limitations, or excessive thrust loading. When this occurs, an increase in the number of propellers is required. Since a single shaft is desirable from an economic viewpoint, there is justification for considering two propellers in tandem on the same shaft. Only small losses in propulsive efficiency (2.2 percent) were reported from model tests in reference [37] for the tandem arrangement over a twin-screw arrangement for a large tanker. The economy of a single propulsion plant, as opposed to two propulsion plants, in addition to the sing%-screw simplicity of the shafting arrangement are the advantages offered by a tandem propeller arrangement. Contrarotating Propellers. Contrarotating propeller arrangements consist of two propellers positioned in tandem on coaxial shafts which rotate in opposite directions. Higher efficiencies can be achieved with this propeller awangement because no rotational energy need be left in the propeller wake. Reference [37] reported a propulsion efficiency improvement of 6.7 percent for a 136,000-ton-displacement tanker with contrarotating propellers as compared with a conventional single-screw arrangement; similar tests for an 18,170-ton-displacement dry cargo ship indicated a 12-percent improvement [38]. Contrarotating propeller arrangements have not been used in connection with commercial ships due to the mechanical complications involved with the coaxial propulsion system arrangement (see Chapter 9). Should the advantages of a contrarotating system become enhanced because of higher horsepower, higher fuel costs,

PROPELLERS, SHAFTING, AND SHAFTING SYSTEM VIBRATION ANALYSIS

MARINE €NGINEERING

h

or lower first costs of the drive system, contrarotating systems may appear ip merchant ship applications. Some naval installations have been made, but their performance has not been made public. FdEy Cwitating Propellers. The primary chjection to propeller cavitation is the deleterious effect that it has on the propeller blade surfaces. Once the propeller loading conditions become such that cavitation can no longer be avoided, as may be the case with very fast ships, then rather than accept a limited amount of cavitation a more satisfactory choice is to design the propeller such that it cavitates fully. In this event, the cloud of vapor which forms on the suction side of the blades does not collapse until it is clear of the propeller blade, thus having no deleterious effect on the propeller blades. Operation at offdesign conditions may result in severe propeller cavitation erosion; and such operation (accelerating, decelerating, etc.) cannot be entirely avoided in service. For this reason and to withstand the high stresses resulting from the large thrust load, fully mvitating propellers are frequently made of exotic materials. Reference [ I ] gives the expected performance of a theoretical series of 3-bladed supercavitating propellers which can be used for estimating performance. In order to achieve fully cavitrvting performance in a speed range too low for the usual fully cavitating propeller design, but still in the range where conventional .propellers would cavitate excessively, ventilation may be conmdered. Ventilation is the term used to describe the introduction of air into the cavitation areas to produce a fully developed cavity. Experience with ventilated propellers is very limited, but some model testing has been carried out; reference 1397 is a report of one such test. 6.3 Propeller Characteristics. An underatanding of the considerations and trade-offs involved in selecting the design characteristics of a propeller is required when developing a shafting arrangement. Assuming that the ship's power and speed requirements have heen established preliminarily as outlined in Chapter 1, the following propeller characteristics must be settled: Propeller Diameter. In general, higher propeller e%iciencies are associated with larger propeller diameters and lower shaft revolutions. Therefore, it is u m d y desirable to install the largest propeller diameter that can be accommodated by the hull structure. The propeller diameter is limited by the lines d the ship as discussed in Section 2.2. PropeUer rpm. The choice of the propeller rpm involves establishing a balance between propeller efficiency and the weight, cost, and space requirements of the main machinery. This is accompliihed by using standard propeller series data (such as reported in reference [l])to compute a series of curva of propulsive efficiency versus revolutions for various propeller diameters. The envelope of these curves indicates the best efficiency and optimum propeller diameter that can be obtained at each propeller rpm. The point of maximum efficiency on this curve for a propeller diameter is termed the

optimum propeller rpm. This envelope curve is used to assess the sacrifice in &ciency which must, be awepeed for any increase of revolutions. Data fmm this curve combined with the effect of the revolutions on the weight, cost, and space requirements of the main p p d l i n g machinery permit the final selection to be made. It will be noted that at revolutions slightly higher than the optimum propeller rprn for a given propellerdiameter, the propeller efficiency d e e m only slightly. But on the other hand, the effect of relatively small incraws of propeller rpm (with the power remaining the same) on the weight, cost, and spme requirements of the main machinery can be significant. I n the case of higherpowered vessels, it is usual to select a propeller rpm higher than optimum and to w p t some sacrifice of propeller dciency in order to reduce the &e of the propelling machinery. Propeller revolutions higher than the optimum are dm accepted for mamns related to the type of maehery. Direct-drive diesel engines and electric-drive machinery usually operate a t speeds higher than the optimum to permit the use of a smaller engine or motor; %veij&t,cost, and space requirements are factors of major importance with these types of machinery. N a d w of Blsdes. Propellers may have t h e , fod, five, six, seven, or more blades. Over the years, the trend has been to use a larger number of blades; three blades fell into complete diswe for large ships during the 1940's. During the I W s , six- and seven-bladed propellera came into use. The major factor in the =leetion of the number of propeller blades is vibration considerations. Both the hull hydrodynamic pressure forms and the forces transmitted through the shafting system bearings are strongly influenced by the selection of the number of propeller blades. I n general, the,propeller exciting forces decrease rapidly with larger numbers of blades; however, there are exceptions. For more detrtib concerning the relationship between We number of propeller blades and the vibratory f o m generated, see Section 3.2. Prudent selection of the number ofpropeller blades iq an important variable which can be used to avoid the excitation of natural frequencies in the propulsion system. PropeEkr Pit&. The selection of propeller pitch can be made when the power, speed, revolutions, and general hull characteristics have been settled. The pitch ratio may be selected on the basis of standard propeller m d e l series data. However, when a propeller is highly loaded or operates in a non-uniform wake field, it may be desirable to design a propeller with a pitch ratio and pitch distribution tailored to suit the particular operating conditions. This can be accomplished by desiping a propeller in accordance with the circulation theory (see reference [I] for a description of circulation theorgr methods). BE& Skew. A propeller blade is termed skewed when its outline is asymmetrical with respect to a s t d g h t radial reference lime inthe plane of the propeller. Skew is usually introduced by successively &placing the

4

1

,

blade sections awav from the direction of rotation. Propeller blades with skew tend to enter and leave the regions of high wake more gradually, resulting in a reduction of the alternating propeller loading due to wake irregularities. The results of model tests have shown that blade skew is an effective technique for reducing the fluctuating forces and moments acting on a propeller. It is normal practice to skew propeller blades a moderate amount based on past experience, without specific knowledge regarding the benefits achieved. Developed Area. With heavily loaded propellers, which is usual with most modern ships, the developed area must be established with care. Considerations in the selection of the propeller developed area are the penalty in efficiency associated with an excessive develo~edarea and the effects of cavitation due to an inadequate developed area. Effects due to inadequate area can be of greater consequence than those due to an excessive area; therefore, prudent practice dictates that a developed area be provided which is sufficiently large to entail a minimal cavitation hazard. kcor a more detailed discussion of propeller cavitation (and consequently developed area) see reference [I]. Propeller Blade Thickness. Requirements concerning the minimum allowable blade thickness are given in classification society rules such as i~ference[18]. A thorough discussion of the development of the classific* tion society rules is given in reference [40], which in addition provides the basis for making an in-depth analysis of the propeller blade stress. Prweller Hub. The controllintz dimensions for the propeiler hub outside diameter an; length are the stern frame (or strut barrel) and the propeller blade fore-andaft length at the interface with the hub. These parameters onlv establish the lower limit. and thicker hubs may be rkquired to provide adequate strength. Excessively large propeller hubs are disadvantageous in that they increase the expense of the propeller and propeller weight (and consequently propeller shaft stress). Propeller Weight. An estimated propeller weight can be obtained in several ways. The most accurate is to conduct a calculatibn based on detailed drawines. Unfortunately, however, the need for the weight has usually passed by the time detailed drawings are available. There are a number of approaches which may be used to approximake propeller weights; one approach is given in reference [40]. One of the less accurate methods, but one requiring the least information, is given in [41J; this method entails the use of wrves which relate the weight of a propeller to its torque rating. There are other methods such as

I

where

a

W

= propeller weight (manganese bronze),

lb (including hub) K = conhnt, approximately equal to 0.26 D = propeller diameter I

MWR

BTF

= mean width ratio

-

developed area per blade D (blade radius - hub radius)

=

blade thickness fraction

maximum blade thickness extrapolated to shaft axis D Care must be exercised in the use of approximate methods because of considerations such as unusual hub dimensions and allowances for ice strengthening. 6.4 Manufacturing Tolerances. As indicated in Table 8, there is an array of tolerances which control the dimensional accuracy of propellers. Table 8 gives a range of propeller tolerances which have been used and also shows recommended tolerances which are expected to produce satisfactory results. Unless care is exercised in establishing manufacturing tolerances, the tolerances can easily be over-specified; that is, the tolerances may be so tight that increased manufacturing costs may be incurred with no corresponding increase in the value of the finished propeller. Various approaches have been used to check the dimensional accuracy of propellers. The usual technique used with merchant propellers is to measure the blade thickness and pitch at discrete points and to check the remaining blade surfaces for fairness. A flexible steel straight edge is held against the blade contours as a means of detecting irregularities of the propeller surface. The technique used with naval propellers is considerably more rigorous; Cylindrical, edge, and fillet gages (which are sheet metal templates machined to the desired blade contour) are prepared which make it possible to comprehensively check the conformance of the propeller blades, at a number of radii, to the design dimensions [42]. In addition to the tolerances governing the propeller physical dimensions, balance tolerances are also specified. Ship's specifications usually require that propellers be balanced (with static or dynamic equipment) such that the static unbalanced force at rated rpm is no greater than one percent of the propeller weight. The following expression may be used to determine the static unbalance corresponding to an unbalanced force equal to one percent of the propeller weight:

where

U

static unbalance which will generate an alternating force equal to one percent of the propeller weight, in.-lb W = propeller weight, lb N = maximum rated propeller rpm =

Limits are not generally placed on dynamic unbalance because of the large diameter-length ratio of propellers, but good practice dictates that corrections made for static unbalance be accomplished so as to improve the

I i

388

MARINE ENGINEERING

389

PROPELLERS, SHARING, AND SHAFTING SYSTEM VIBRATION ANALYSIS

I

Table 8 Dimension Blade Thicktw.88 Maximum Minimum

Blade Width Maximum Minimum

Propeller Manufacturing Tolerances

-

Naval

Rmge of Experience Merchant -0.00, +% in. or 2% of deaign th~ekness,whichever is greater -0.00, +% in.

Recommended Practice

+%.$a in. or + l % of design -0.00 +>$ in. or 2% of thickness, whichever is greater deaign thickness, whichever is greater - .02 in. % % of thickness) d 4 in. +I % of thickness)

+

+

+g2in. or 1% of width 4

in. +0.25% of width) in. +0.5% width)

+>b in.

iRd2

Bhde Position in Transverse Phne +15 min Maximum +30 min +15 min &15 min Minimum Blade T r d at Tip +1.0 in. Maxlrnum &%6 in. & x 6 in. Minimum +% in. Pro e* Diameter -1.25 in., +0.00 daxlmum &% in. Minimum +% in. -.20 in., +O.OO Deviation of Pitch cf.tany Radius from Designed value Maxlmum +2% +1.5% Minimum % +l.O% Geatest Variation i Any Avetage Blade Pitch to Average Propeller Pitch .Gum +I% Mi imum &0.75% z?5%

dynamic unbalance. Dynamic unbalance is generally not found to be a problem; nevertheless, the dynamic unbalance should be limited such that the alternating force generated at the aftermost bearing is no greater than an alternating force a t the aftermost bearing corresponding to a static unbalance equal to one percent of the propeller weight at the propeller center of gravity. A useful expression for the maximum allowable dynamic unbalance under these conditions is as follows:

+15 min

&% in. &% in.

0

where

D

= dynamic unbalance which will produce same

force at aftermost bearing as a static' unbalance equal to one percent of propeller weight applied a t propeller center of gravity, in.-lb-in. W = propeller weight, lb L1 = distance from propeller center of gravity to aftermost bearing reaction, in. Lz = distance from aftermost bearing reaction to reaction of next bearing forward, in. N = maximum rated propeller rpm E = shaft modulus of elasticity, psi I = shaft rectangular moment of inertia, in.4

Section 7 Torsional Vibration 7.1 General. Severe torsional vibration difficulties experienced with the early reciprocating engine drives and particularly diesel engines moved the importance of torsional vibration as a design consideration to the forefront. Subsequently, torsional vibratian became established as a factor which had to be carefully considered in the design of all types of main propulsion shafting systems, and the design methodology required to conduct a reasonably accurate torsional vibration analysis has been formulated [41,4349]. A summary of the analytical methods used to evaluate the torsional vibration characteristics of geared turbine-

driven main propulsion shafting systems is included in the following pages. The fundamental theory of torsional vibration is well documented in the existing literature; therefore, it will not be reproduced here. A discussion of details peculiar t o torsional vibration of diesel-driven main propulsion shafting systems is similarly not included. 7.2 Modes of Torsional Vibration. The design of most large turbine-driven ships is such that one or more resonant modes of tonional vibration will occur within the operating range. The first mode of torsional vibration is the one in which the node is immediately abaft

the slow-speed gear and the antinode is a t the propeller; in this mode, the inertia of the propeller vibrates against the inertia of the turbines and gears. For vessels with fairly long runs of shafting, the first-mode frequency is excited by blade rate excitation at a very low propeller rpm (at about % of maximum rated rpm). For this reason it is seldom objectionable as the alternating torques developed are not of sufficient magnitude to be deleterious. On the other hand, in the case of ships which have short runs of shafting, the first mode may occur above 50 percent of the maximum propeller speed and warrant a comprehensive analysis. Second and third torsional vibration modes are determined primarily by the characteristics of the prime mover. With geared turbine drives, the turbine-gear system generally cannot be designed such that the second mode of torsional vibration (the one in which the two turbine branches vibrate against each other) is out of the operating range. This being the case, a so-called "nodal drive" is frequently provided [43]. In a nodal drive, the turbine branches are designed to have equal frequencies; this forces the slow-speed gear to be a nodal point. The second mode of torsional vibration will then consist of a motion in which the two turbines vibrate so that their vibratory moments oppose each other with a nodal point a t the gear. This being the case, the turbine branches cannot be excited by the propeller. In the third mode of torsional vibratiqn, the vibratory torques of the propeller and turbines oppose that of the slow-speed gear. The third mode usually occurs considerably above the operating range; consequently, it is rarely of concern. However, very high rpms or a large number of propeller blades may bring it within the operating range. The third mode is difficult to excite because the antinode occurs a t the slow-meed gear, which is not a source of excitation with modern gears, and the node occurs near the propeller, which is a source of excitation but has a very small vibratory amplitude. The mode shapes of the first three modes of torsional vibration are shown in Fig. 17. 7.3 Models for Torsional Vibration Analyses. A typical steam turbine propulsion system is schematically illustrated in Fig. 18(a). From an .inspection of2Fig. 18(a), it is apparent that a comprehensive torsional vibration analysis of such a system would be prohibitively complex if the classical approach were used; cobequently, simplifications must be made to facilitate a practical analysis. The system can be reduced to an equivalent svstem (model) in which all elements are referred to the sime rdtationil speed, thereby greatly simplifying the analysis. Such a procedure may be used to obtain the model shown in Fig. 18(b). Figure 18(b) can be used to evaluate all modes of torsional vibration which would be expected to be of interest in practice. However, if only the first three modes of vibration are of interest, which would generally be the case, the model shown in Fig 18(b) can be further simplified to that shown in Fig. 18(c) without a serious loss of accuracy due to the fact that the equivalent inertia of the turbines and stiff-

( a ) First Mode

( b ) Second Mode

Legend I

2 3 4 5 6

L P Turbine L P High-Speed Reduction HP Turbine H P High-Speed Reduction Slow-Speed Retjuction

Propeller

5

(c) T h i r d Mode fig. 17 Mode shapes of first three modes of tonional vibration of a turbinedriven propulsion system of nodal-drive type

ness of the turbine shafts are very high compared with those of the first reduction gear elements. If only the first mode of torsional vibration is of interest, then it can be approximated in either of two ways. One way would be to directly add the equivalent inertia of the turbine branches [JTL and JTH in Fig. 18(c)] to the slow-speed gear inertia (Ja) and make an analysis based on a two-mass system. A more practicable approach, since the inertias of the turbines and gears are . frequently unavailable, is to assume that the nodal point in the first mode of vibration is four percent of the distance from the slow-speed gear to the propeller aft of the slow-speed gear. With such an assumption, the first-mode natural frequency can be simply determined by considering the system to be a one-degree-of-freedom model as shown in Fig. 18(d). All of the system parameters which are needed in order to evaluate the torsional natural frequencies can be directly determined from the physical properties of the system except for the propeller entrained water. Assessment of the propeller entrained water can be made from the work of Burrill and Robson [49]. I n order to avoid the tedious labor associated with calculating the moment of inertia of the propeller in air, an approximate method such as that described in reference [41] may be used or the propeller radius of gyration may simply be estimated to be between 0.40 and 0.44 of the propeller radius (lower values correspond to larger propeller hubs and smaller numbers of blades). 7.4 Determination of Natural Frequencies. The Holzer method of computing the natural frequencies of lumped spring-mass systems is a convenient procedure for determining the torsional natural frequencies of

MARINE ENGINEERING :ond reduction

PROPELLERS, ,SHAFTING, AND SHAFTING SYSTEM VIBRATION ANALYSIS Table 9

rn

LP f i r s t HP f i r s t

Determination of First Natural Mode of Torsional Vibration for a Turbine Driven Propulsion System Modeled as Shown by Fig. 18(b) Number of pro eller blades = 6 Assumed propefler rpm = 22.38 J = in.-lb-set' + 1W k = in.-lb/rad + 10"

reduction r o t i o = n 2 reduction r a t i o = n,

391

o =

oa =

6(22.38) (2a)/BO rad/sec 197.7 rad'/sec"

( a ) S c h e m a t i c i l l u s t r o t i o n of a geored t u r b i n e d r i v e n propulsion system

B L = -0.06477 Bp 011 = -0.06417 Bp 2Q = 1213 B L 321 .O OH 98.8 BP 2Q = -78.57 Bp 20.60 BP 98:s BP

+-

++

2Q = 0; therefore a resonant condltlon

( b ) Equivolent 6 moss system w i t h a l l bronches r e f e r r e d to t h e propeller r p m

(c) Four mass system w i t h o l l bronches r e f e r r e d t o t h e p r o p e l l e r r p m f o r opproximoting t h e f i r s t t h r e e n o t u r o l frequencies o f torsionol vibrotion

Fig. 1 8

( d l Single degree o f freedom system f o r opproximating t h e f i r s t noturol frequency of torsionol vibrotion

details concerning torque excitation, see references [3, 50-521. 7.6 Damping. There are several sources of damping which tend to reduce the maximum attainable amplitude of torsional vibration; one of the most important is the propeller, particularly in the first mode due to its being a t the antinode. Propeller damping can be determined in several different ways [44-46, 531. I n general, the propeller damping coefficient, b, can be expressed as

Equivalent systems for determining natural torsional f requencies of geared turbine-driven propulsionsystems

b = KQ/Q

in.-lb-sec rad

where turbine-driven propulsion systems. I n order to review the computational procedure for a practical example, consider a turbine-driven vessel that is modeled as shown by Fig. 18(b). Typical values of system inertia and spring constants and calculations for the first torsional natural frequency are given in Table 9. The calculation is accomplished by first assuming the resonant frequency of the system, relating this frequency to the propeller rpm, and then calculating the corresponding vibratory torque and torsional amplitude a t the slow-speed gear (inertia JG) in terms of the amplitude a t the terminal end of each branch. For convenience the amplitudes a t the terminal ends of the three branches are initially assumed to be one radian. Since the three branches (propeller, LP turbine, and HP turbine) must have the same amplitude a t the slow-speed gear mass, the amplitudes of the three branches can be expressed as a function of the same unknown amplitudefor instance, the propelle~thereby obtaining the mode shape. The torques imposed on the slow-speed gear are then summed; if the sum is zero, a resonant condition is established. If the sum is not zero, the process is iterated until the sum is zero by assuming a different resonant frequency. The same procedure can be repeated to determine the remaining four torsional natural frequencies but, as

previously mentioned, only the first three modes would generally be of interest. The mode shapes of the first three natural torsional frequencies are shown in Fig. 17. The node in the first mode is seen to be immediately abaft the slow-speed gear. The two turbine branches are tuned in the second mode such that the slow-speed gear is a nodal point. The third mode is the one in which the slow-speed gear is the antinode with the terminal end of the three branches being near nodal points. 7.5 Excitation Factors. There are several possible sources of torsional vibration excitation with turbinedriven ships, but the propeller is the only one of consequence. With the accurate cutting of modern gears, gear-excited criticals are either wholly absent or are of negligible amplitude. Propeller excitation of a frequency higher than blade rate exists but it is normally negligible in magnitude. A number of factors, such as the propeller loading, propeller aperture clearances, number of propeller blades, hull lines, and hull draft, influence the magnitude of the vibratory torque; consequently, generalizations in this area must be used with care. Nevertheless, typical ranges of torque excitation, expressed as a percentage of the mean torque, are presented in Table 2; for additional

K = a constant Q = mean propeller torque, in.-lb Q = rotative speed of propeller, rad/sec

If propeller model test data are available, it may be shown that

dependent and independent variables, moving along a constant pitch line. The energy loss via the propeller per cycle of torsional vibration can be written as

E,

= ~ b w 8 , ~ in.-lb

(201

where w = circular frequency of vibration, rad/sec 8, = amplitude of propeller vibration, radians

Energy is also dissipated as a result of elastic hysteresis in the shafting, sliding fits, etc. Although such internal damping losses in a shafting system would be expected to be small when considered individually, they are frequently estimated to dissipate about 5 percent of the vibratory energy per cycle when totaled. The energy dissipated due to internal damping crtn be expressed as [46]: I

where where s is the propeller slip. If the propeller data are given in the form of J, KQ curves

If given in the form of a Troost diagram (B,, 6) as in reference [I] :

As an approximation for many propellers, K = 3.7 to 4, which may be used in the absence of other data; a value of 4 corresponds to a damping constant which is double the slope of the torque-speed curve. I n all cases the derivatives are computed a t the operating point of the propeller by taking the ratio of small differences in

ar = fraction of energy dissipated; the value of this

quantity is approximately 0.05

J,

= moment of inertia of mass n, lb-in.-sec2

8, = amplitude of vibration of mass n, radians

9

The damping action of the turbines would general be expected to be of secondary importance especial1 in modes where the turbines have small relative amplitudes; however, it may warrant assessment under some circumstances. The energy dissipated due to turbine damping can be expressed as

ET =

z ~ c w 8 , ~ - in.-lb

(22)

where c = turbine damping constant which can be approximated as the ratio of turbine torque to turbine rpm a t the speed corresponding to the point under study, in.-lb-sec/rad

392

MARINE ENGINEERING

PROPELLERS, SHARING, AND SHAFTING SYSTEM VIBRATION ANALYSIS

w = frequency of vibration, rad/sec 8, = amplitude of vibration of turbine rotor, radians

If damping is introduced into the vibration calcolrlr tions, the computational procedure is modified considerably. An external source of damping, such as that at the propeller or turbines, introduces an external moment of -jbw8 on the respective mass concentration; and internal damping, such as shafting hysteresis, between two masses is equivalent to changing a spring constant k to a complex spring constant

evaluation. These attempts have not proven successful due to many variables which must be considered. For example, the propeller design (pitch distribution, skew, number of blades, etc.), propeller aperture clearances, hull lines in the vicinity of the propeller, harmonic content of the wake in way of the propeller in conjunction with the type of reduction gear, system damping, ship operating practices (shallow water, partial propeller emersion, operating point) and similar considerations vary so much from one ship to the next that there are numerous exceptions which can be taken to all, except the most trite, generalizations.

and by substituting the expression 4Q/Q, equation (16), for the propeller damping coefficient and letting w = ZQ, equation (25) can be simplified to

where Z is the number of propeller blades. The maximum amplitude of propeller vibration can be determined from the foregoing expression. In conjunction with the normalized mode shape determined from the system natural frequency calculation, Table 9, where ar is the fraction of the elastic energy absorbed by the propeller amplitude is used to assess the vibratory torque at resonance at any element of the system. The the damper [45]. Calculations which incorporate damping as just alternating torques in the quill shafts between the highindi'cated are somewhat tedious, particularly if the cal- speed gears and low-speed pinions are usually the largest culations must be done by hand; an easier procedure is to from a relative viewpoint; consequently, it is customary compute the effect of damping at resonance only by to analyze these elements when investigating the possibility of torque reversals. equating input energy to dampening energy. As an example, referring to the calculation in Table 9, 7.7 Vibratory Torque Calculations. In many cases, with a propeller excitation equal to 3 percent of the mean the torsional vibration characteristics of a shafting propeller torque, the alternating torque, q, in the lowsystem can be shown to be satisfactory in the design stage with only a computation of the system natural pressure quill shaft in the first mode of torsional vibration frequencies and without predicting vibratory torques would be : q = 1213 8~ X lo6 and amplitudes. Normally this is possible when a q = 1213 (0.06477 8,) X lo6 compar&on is made with a similar system that has q = 78.6 (r/4Z) X lo6 proven satisfactory in service. For designs where the q = 98,300 in.-lb system natural frequencies, vibratory excitation, or anticipated system operation may cause concern, investi- This is the torque in the low-pressure quill shaft referred gation of the magnitude of the vibratory torques and to line shaft speed. With a second reduction ratio of stresses is necessary. 7.5, the actual vibratory torque in the quill shaft will be In order to illustrate the procedures used to assess 98,300/7.5 or 13,100in.-lb. In this particular case at the vibratory torque amplitudes, again consider the calcula- resonant frequency, the low-pressure turbine develops tions shown in Table 9. In the absence of speed-power 55 percent of the total power delivered to the propeller. curves for the ship, the mean operatingtorque correspond- The mean torque in the low-pressure branch a t resonance ing to the resonant frequency can be approximated by is consequently 33,500 in.-lb whereas the alternating determining the rated propeller torque (the torque torque is estimated to be 13,100 in.-lb; therefore, torque corresponding to 22,000 shp at 115 rpm) and assuming reversals in the low-pressure train at the first resonant that the propeller torque varies as the square of the mode of torsional vibration are not expected. The propeller rpm; therefore, the mean operating torque at vibratory stress in the quill shaft, kz, can be calculated the resonant frequency is estimated to be using the alternating torque across this shaft of 13,100 in.-lb. In the more general case, where system damping in addition to that associated with the propeller is of importance, the maximum amplitude of propeller vibraThe exciting torque can be expressed as rQp where tion at resonance can be found by solving the following taken from Table 2, is the alternating torque expressed equation for 8p: as a fraction of the mean torque. With a maximum , vibrating energy propeller exciting torque of r ~ the EE = E, El ET (27) input per cycle of vibration is: 1 nrQ8, = ~ b ~ 8 ; 5 T C U B ~ ~ (28) Y EE = nrQ8, (24) All of the terms, except 8,) are either known or can be In the first mode of vibration, the propeller would be expressed as a function of 8,. Once 8, is established, the expected to be the only significant source of excitation calculation can be continued as shown. and damping; therefore, by equating the expressions for 7.8 Acceptable Limits for Torsional Vibration. the propeller excitation and damping energy, the maxi- Many attempts have been made to standardize the mum propeller amplitude can be determined as follows: procedure to be used for torsional vibration analysis and

+ + +C

+C

393

As a broad rule it may be stated that untuned torsional vibration resonant frequencies should not occur in the range of 60 to 115 percent of rated rpm; however, this rule does not in fact ensure satisfactory torsional vibrs tion characteristics; furthermore, there may be aatisfactory systems in service which it would exclude. Generalizations can serve as a broad guide, but the hazard associated with generaliiations is that their limitations may not be appreciated. There appears to be no satisfactory alternative to conducting an analysis of each particular system and studying each factor individually.

Section 8

I

8.1 Introduction. Severely objectionable longitudinal vibrations were not encountered until the advent of several classes of large naval vessels in early 1941. Reference [54] contains a description of the difficulties experienced with these ships and also presents the most thorough treatment of longitudinal vibration that has been prepared. The works of Panagopulos [47], Rigby [55], Couchman [56],and others have added to the knowledge of the subject; yet the fundamental problem areas encountered today are the same as those identified by Kane and McGoldrick [54]. The low level of shaft axial stress associated with even the most violent instance of longitudinal vibration is not sufficiently large to induce failures in the shafting itself; nevertheless, longitudinal vibration can produce effects which are destructive to engine room equipment. Shafting systems which have longitudinal vibration characteristics that are resonant with propeller blade rate f r e quency forces experience a significant magnification of the exciting forces. Such a force magnification can result in such deleterious effects as:

1 Accelerated wear of gears, flexible couplings, thrust bearings, etc., and destruction of turbine clearances due to the increased relative axial movements. 2 Large vibration amplitudes and stresses in steam piping, condensers, and main and auxiliary machinery which ultimately result in fatigue failure. 3 Cracks in foundation and hull structures. 8.2

Determination of Natural Frequencies. There

are basically three approaches which may be taken in determining the natural frequencies of longitudinal vibration. The first approach would be to use a simplified method for the purpose of quickly assessing a situation. Approximate methods suitable for investigating the firstmode of vibration are given in references [47,,541. A second approach would be to model the shafting system as discrete masses and springs and use the Holzer method to determine the system natural frequencies.

The accuracy obtained with a discrete spring-mass model will depend upon the masses used to represent the shafting system. A third approach is the mechanical impedance method proposed by Kane and McGoldrick [MI. This method is inherently more accurate than the Holzer method since the weight of the shafting is considered to be distributed; however, the impedance method has the slight disadvantage of being somewhat more complex and diicult to grasp. For illustrative purposes, a calculation of the natural modes of longitudinal vibration of the shafting arrangement shown in Fig. 1 will be made using the mechanical impedance method. Figure 19 is a model of the shafting arrangement which is suitable for analysis by the mechanical impedance method. I t may be noted that the difference in the diameter of the inboard and outboard shafting is taken into account; in general,

M p ,= mass o f p r o p e l l e r , 227 lb-sec2/in M g = mgss c f gears, 147 lb-sec2/in M c = mass o f machinery, 7 6 7 lb-sec2/in m l = mass o f larger (outboard) shafting, 301 lb-sec2/in m 2 = mass o f s m a l l e r (inboard) s h a f t i n g , 3 6 6 lb-sec2/in

kt

= spring constant of l a r g e r (outboard) shafting,

k2

2 0 . 3 x 10' Ib/in = spring constont o f smal ler (inboord) shafting, , 5.9 x 10' lb/in

= t h r u s t bearing spring constant, 7 x 10' w i n

kt kf

= t h r u s t bearing foundation spring constant, Ib/in

2

' number of

p r o p @ l l e r blades, 6

Fig. 19 Representation of a geared tutbina propulnion hafting system ( ~ g I. ) for a longitudinalvibration analysis

394

PROPELLERS, SHAFTING, AND SHAFTfNG SYSTEM VIBRATION ANALYSIS

MARINE ENGINEERING

especially with short spans of shafting, this additional degree of sophistication is not yarranted. The majority of the system parameters may be directly calculated from the system scantlings, and therefore no difficulty is experienced in obtaining their value. However, assessment of several of the system parameters can be nebulous. For example, determination of the water entrained with the propeller does not lend itself to an accurate calculation; as a first approximation the entrained water weight may be assumed equal to 60 percent of the propeller weight. The results obtained from the experimental work of Burrill and Robson [49] are widely used in estimating propeller entrained water, and reference [54] suggests other approaches. The behavior of flexible couplings in connection with vibratory movements similarly cannot be stated with certainty. The impact of the behavior of flexible couplings, aside from the effect on the couplings themselves, is not great when the thrust bearing is located well forward. But when the thrust bearing is located aft and there is an appreciable vibratory amplitude a t the slow-speed gear, the impact of flexible coupling behavior can be significant. For a detailed discussion of the behavior of flexible couplings and the complications involved, see [54 and 561. Some machinery liquid and foundation weight participates with the shafting system when vibrating longitudinally as a consequence of being near the main thrust bearing; but assigning a magnitude to these quantities entails numerous uncertainties. Assessment of the "machinery mass," M,, to be included in the mathematics requires judgment which must be based on the specifics of each system. In instances w-here only the first mode is of importance, the machinery mass has a small participation and consequently an accurate assessment of its magnitude is not critical. On the other hand, the machinery mass is expected to have a significant participation in the second mode, in which case care must be taken in its determination. In general, the first reduction gear rotating elements, gear casing, turbines, condenser, foundation structure, or portions thereof may be included as machinery mass. Reference [54] gives some guidance in the assessment of the machinery mass. Reference [57] contains an interesting approach on the treatment of machinery masses in that a portion of the machinery mass is given a leverage ratio relative to the centerline of the shaft. The spring constant of the thrust bearing, k t b , may be considered to consist of three constituents: the spring constants of the thrust bearing housing, the thrust collar, and the thrust elements (or shoes). Aside from the tedious calculations, no diiculty is experienced in calculating the spring constants of the thrust bearing housing and collar inasmuch as the majority of the ' deflections are due to shear and can be estimated on the basis of well-established techniques. But the spring constant of the thrust elements can be difficult to evaluate. In the absence of more specific data, the data

given in [54] may be used for guidance in establishing the stiffness of thrust elements. Determination of the thrust bearing foundation spring corlstant can be a difficult and nebulous undertaking even for an experienced analyst. In general, the thrust bearing foundatiori structure is arranged such that an accurate calculation of its spring constant would be formidably complex. For this reason, longitudinal vibration calculatior~sare frequently conducted such that the natural frequency is expressed in terms of the thrust bearing foundation stiffness. Table 10 contains a calculation for the first and second resonant modes of longitudinal vibration of the shafting system modeled as shown in Fig. 19. Table 10 utilizes the mechanical impedance method described in [54] and is arranged such that the resonant frequencies can be plotted in terms of the thrust bearing fourldation stiffness. Figure 20 is such a plot and indicates the accuracy required of the thrust bearing foundation stiffness calculation. In some instances, as may be the case with a ship having a very short run of shafting, an inspection of the appropriate thrust bearing foundation drawings may be all that is required to provide assurance that the resonant modes of longitudinal vibration will be well clear of the operating range. On the other hand, lengthy and sophisticated thrust bearing foundation stiffness calculations may be required in order to ensure that ships with long runs of shafting have satisfactory longitudinal vibration characteristics. The thrust bearing foundation spring constant may be considered to corlsist of three constituents: the bending deflection of the thrust bearing foundation structure above the inner bottom, the shear deflection of the thrust bearing foundation structure above the inner bottom, and the deflection at the thrust bearing due to inner-bottom deflection. An appreciation for the effects of innerbottom deflection may be obtained from reference [55]. Calculation of foundation deflections above the inner bottom entails the usual obscurities associated with estimating the deflection of complicated structures. In ' order to assess the effects of inner-bottom deflection, it is necessary to make simplifying assumptions concerr~irlg the extent (length and breadth) of inner bottom effectivelj. supporting the thrust bearing and the boundary conditions at its extremeties; each system must be individually studied, in light of the degree of accuracy desired, in order to establish appropriate assumptions and calculation procedures. In cases where design constraints make it impossible to design a shafting system such that it is free of objectionable frequencies of resonant longitudinal vibration, use may be made of a "resonance changer. " Resonance changers are discussed in Chapter 20. Briefly, they are thrust bearings which are modified such that the thrust pads are floating pistons. The volume of oil supporting the thrust pads can then be tuned to alter the thrust bearing spring constant and avoid objectionable resonant frequencies.

Longitudinal Vibration Calculations for Shafting System Model Shown by Fig. 19

Table 1 0

N

o0

rpm

70 80 90 95

US

red/sec 43.98 50.27 56.55 59.69

3

395

x 106

Z,, = osMp ib/in. X 106

0.0193 0.0253 0.0320 0.0356

0.4381 0.5743 0.7264 0.8081

tan a1° = Zb/kl€l

kla tan a1°

"=

180

VulR1, e. The Head-Capacity Curve, Effect of Shape of and the impeller exerts a net torque on the fluid, the Impeller Vanes. In the absence of stationary guide vanes total theoretical torque is at the inlet, a desirable simplification is obtained by assuming that the absolute fluid velocity at the inlet is radial. Thus, there is no peripheral component regardless of the shape of the impeller inlet (that is, Vul is zero). (8) Sincse Vl is perpendicular to ul, it follows that v12 - V12 - u12 equals zero because the three terms are respecIn the foregoing, Vu2R2and VulR1are assumed to be con- tively the hypotenuse and adjacent sides of a right tristant over the two surfaces of integration. angle. The expressions previously given for the theoretiSince Vldal and V2da2are equal and represent the total cal head become rate of flow through the impeller, then Q = volume rate of flow, cu ft/sec q = volume rate of flow, gpm T = theoretical rotor torque, ft-lb H , = theoretical pump head, ft , a = acceleration due to gravity, ft/sec2 -

+

+

+

's"

II

The expressioti for work is obtained by multiplying the impeller torque by the angular velocity o. After dividing through by the weight rate of flow, there is obtained the usual equation foi the theoretical head: 1 (10) 9 d. Physical Interpretation of Theoretical Head. The foregoing ex~ressionfor the theoretical head may be traniforked b y means of the cosine law. ~ e f e r r i i gto

Ht

=

To

--

w

=

- (ulVu2- ulVul)

Numbers in brackets designate References at end of chapter.

LEGEND TUEDRETICAL HEAD

To show the relation of theoretical head to the volume flow, or capacity Q, it is necessary to make another application of the cosine law, whereby 9 The im~ellerexit area Az (in the diredtion of vz) is rD2B2 sin '& in square feet (neglecting the vane tip?, and 02 equals Q/Az, SO that

3.6

ACTUAL HEAD

I- CIRCULATING PUMP, RADIAL-FLOW TYPE (a) NORMAL CHARACTERISTIC,& =2?.s0

3.4

3.2

+ 3.0

j

2.8

5

2.6

P

2-FEED PUMP. RADIAL-FIDW

TYPE

STEADILY RISING CHARACTERISTIC,&~I~~ 3-DREDGE PUMP, RADIAL FLOW T Y P E , ~ ~ = ~ O ~ 4-CIRCULATING PUMP, AXIAL-FLOW TYPE AXIAL-FLOW CHARACTERISTIC,~m=240

2.4

a

K 2.2

t*

2.0

a

1.8 \

2

1.6

FLOW RATIO

=

FLOWIRATED FLOW

Fig. 10 Theoretical and actual charactariatics of various centrifugal pumps

This equation shows that for a constant rotational speed and with a discharge angle & less than 90 degrees (backward-curving vanes), the theoretical head decreases as the capacity is increased, following a straight line of downward slope. This is illustrated by curve l(a) in Fig. 10. In this figure the head-capacity characteristics are shown in ratio form, with all heads, both theoretical and actual, being divided by the actual head at rated capacity and plotted against the corresponding capacity divided by rated capacity. If the discharge angle is 90 degrees (radial vanes), tan & in the foregoing equation is infinite, the second term of the equation is zero, and the theoretical head is a horisontal straight line independent of the volume as shown by curve l(b). Likewise, if the discharge angle is made greater than 90 degrees (forward-curving blades), then tan /32 is negative and the theoretical head increases with capacity as shown on curve 1(c) . The actual head-capacity characteristic is always less than the theoretical, and the various curves of Fig. 10 reveal the wide discrepancy between the theoretical and actual heads. In the region of the designed capacity, the ratio of actual to theoretical head may be anywhere from

0.25 to 0.75 and this ratio is usually smaller when the capacity is greater or less than the design value. This discrepancy cannot be explained by the elementary theory previously outlined, which necessarily assumes an infinite number of guide vanes of zero thickness. However, the '(circulation theory" of hydrodynamics shows that for a finite number of vanes there must exist a circulation or eddy flow within the impeller which has the effect of reducing the mean peripheral component of the absolute exit velocity [2]. This has been confirmed by experimental work. Curveq 2 and 3 depict the characteristics of other radial pump types. The characteristics of the axial-flow type of pump are shown by curve 4. The theoretical headcapacity curve of an axial-flow pump rises steeply, while the actual head-capacity curve has an irregular shape, being relatively flat near design capacity but rising sharply at lower capacity and falling rapidly at higher capacity. The elementary theory for this type of pump is the same as for the radial type, but as shown in Fig. 11 there is a different velocity diagram for every radius at the pump suction and discharge. Thus, in setting up an equation for the theoretical head, it is necessary to use the velocity diagrams at a mean discharge diameter Dm.

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

MARINE ENGINEERING

11

1.

I

CYLINDRICAL SECTION NEAR RIM. R,

approach is one in which actual test results are used in the development of various correlations. The applic* tion of dimensional analysis that follows is a logical derivation of the various centrifugal pump design constants. Primary interest is focused on the volume flow Q, in cubic feet per second, and the actual net head H, in feet of fluid, of a centrifugal pump, but it is also recognized that the most important physical characteristics of a pump are the speed w, in radians per second, and the impeller discharge diameter Dl in feet. Other physical characteristics, such as inlet diameter, width of impeller passages, inlet and discharge pipe diameters, and the corresponding area. or surfaces are described by ratios of these quantities to the basic diameter D or D2,depending on the number of linear dimensions involved. Two physical properties of the fluid may enter into consideration. These are the weight density p, in pounds per cubic foot, and the kinematic viscosity v, in feet squared per second. The acceleration of gravity g, in feet per second per second, must also be taken into account. Thus the problem contains seven distinct physical quantities, all expressed in three fundamental unitslength in feet, force in pounds, and time inxeconds. The basic theorem of dimensional reasoning holds that a general relation between all seven physical quantities may be expressed in seven less three, or four compound quantities, each of zero dimensions. One general expression of this form is

CYLINDRICAL SECTION NEAR ,HUB, %

I a1 "RI

Fig. 1 1

(1) Impeller inlet vane shock losses. (When diiuser vanes are fitted in the discharge casing, there are also similar losses a t the inlet to these vanes.) (2) Impeller and diffuser vane exit losses, due to eddies formed by the edges of the vanes. (3) Friction losses in the inlet section, impeller, diffuser, and diiharge casing; similar to the friction losses in piping. (4) Additional eddy and turbulence losses in the impeller and discharge casing where kinetic energy is converted into pressure energy. I

I

charge casing back to the inlet through the impeller running clearances. (7) Balance-device leakage where the rotor axial thrust is equalized by a balancing drum or disk. (8) Friction losses in bearings and stuffing boxes, including thrust bearings. The foregoing losses can be calculated approximately from appropriate special theory for each type of loss, and the net pump efficiency determined accordingly. However, it is generally more satisfactory to correlate actual test efficiencies of centrifugal pumps from the laws of similarity; this method of approach is developed in the following. The water horsepower, Pw,for a pump is given by and for a brake horsepo.cver, P, the phmp efficiency, El equals Pw/Por

1" the foregoing, Q is the volume flow in Cu ft/~ec. If q represents the volume flow rate in gpm and p'the head in 1. psi, the efficiency becomes

E = - Pq All of the above losses affect the head generated and (17) 1714P therefore contribute greatly (but not exclusively) to lower the value of the actual head. The following losses affect Additional detsils regarding pump design only the power input: tions are contained in references [3] through [a]. 1.2 Laws of Similitude. Purely theoretical reasoning (5) Frictional losses a t the exterior surfaces of the impeller, similar to windage losses of steam turbine wheels. does not at present afford a very satisfactory basis for the (6) Flow losses due to leakage of fluid from the dis- study of centrifugal pump performance. A more direct

equal -to rD. Thus, it is seen that this last group of terms has the usual Reynolds number (R,) form, except that it is inverted. This group will therefore be of importance when considering the effects of frictional forces on the pump performance, but usually it is secondary in importance to the more general characteristics indicated by the first three dimensionless groups. Since the basic groups in equation (19) are dimensionless, they may be combined without a loss of generality. Thus, multiplying the two middle groups gives gH/ n2DW and: denoting the head coefficient by #,

The first group in equation (19) is the specific capacity, q,, divided by r, that is,

This expression for specific capacity states the wellknown relationship that, for a pump of given diameter, Q varies directly as the speed n. Or, for similar pumps a t a constant speed, Q varies directly as the cube of the diameter. To be correct, the head and flow rules as expressed by equations (20) and (21) must be taken together, and, therefore, if the capacity changes in proportion to the speed, then the head changes in proportion to the square of the speed. Likewise, for geometrically similar pumps, if the capacity changes in proportion to the cube of the diameter a t a constant speed, then the head changes in proportion to the square of the diameter. The specific speed for a particular pump may be deThe requirement that each group must have zero fined as the speed required by a pump of similar design, dimensions makes it possible to determine the values of shape, and hydraulic characteristics to develop a head of the exponents a, b, 6, etc., in the foregoing expression. 1 foot when delivering a t a volume rate of 1 cubic foot This is done by substitution of the dimensional formula per second. Thiis, if a set of values D, n, H, and Q are for each physical quantity in the groups. For example, selected for a given pump (usually the rated values), the first group must then have the form then, in order to have similar characteristics, a second pump must be such that its corresponding physical quantities will give numerical values for all the basic dimensionless groups equal to those obtained with the original pump. hence If Dl, nl, HI, and Q1 represent the physical quantities of the second pump, then the following relations satisfy the requirements for equality of all dimensionless groups:

"RZ

Velacity diagrams for an axial-flow pump

It is seen in Fig. 11 that the development of a cylindrical section through the impeller is a series of vanes resembling air-foil sections. This is the basis for the calculation of axial-flowpump performance from airfoil theory, which has given results that are surprisingly close to the actual pump characteristics. f. Pump Losses and Efficiency. Flattening of the discharge angle, as previously described, causes actual flow conditions at the impeller outlet different from those assumed in the foregoing elementary theory, but it should not be inferred that the difference between the actual and ideal theoretical heads represents a corresponding loss in efficiency. The principal losses in centrifugal pumps are as follows:

409

E

Thus, the first group in equation (18) is found to be Q/wD8, and when the same process is carried out for the remaining groups, with n-n substituted for w, the general expression becomes

The last term in equation (19) may also be written v/uD = l/R,, since the impeller peripheral velocity u is

However, Q1 = H1 = 1 by definition. Combining all three equations:

k

f

I.# 1.6

0

g 1.4 D:

5 3

1.2 1.0

W

50 0.8 $

0.6

22 0.4 2 0.2 %

Y

o

0 0.2 0.4 0.6 0.8 1.0 1.2 1.4 1.6 FLOW RATIO r FLOW / RATED FLOW

Fie. 12

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

MARINE ENGINEERING

41 0

Copacitphead and efficiencycharacteristics of a cedtrifugal pump at various speeds

QIHl -= r8nl8Dl6

1 - QH a8nlaD16 . aanaD5

or

-

multiplying it by the term n/(2g)'/'. Specific speed is the most important single index to the potential performance characteristics of a centrifugal pump. After some limited experience one can examine the shape of a centrifugal pump impeller and make a fair estimate of its specific speed. Then, from curves such as Fig. 14, the efficiency and other performance characteristics of the pump can be estimated. Of course, the brake horsepower , level is also an important index. One great value of the parameter specific speed is that it expresses the requirements for similarity of flow conditions in terms of the three essential pump characteristics-capacity, head, and speed-independently of any physical dimensions ~f the pump. a. Performance Curves for a Single Pump. By means of the dimensionless constants derived in the foregoing, it is ,possible to extrapolate the results of a constant-speed test of a particular centrifugal pump to predict its performance at other speeds and also, within reasonable limits, to predict the performance of any geometrically similar pump of different size. The extrapolation process for a single pump is illustrated in Fig. 12. The solid curve represent, in ratio form, the capacity-head and capacity-efficiency characteristics from actual tests at a constant speed N. To predict the characteristics at a different speed NI, select any point (Q, H) on the test curve and calculate new values (Q1, HI) for the new speed such that the dimensionless constants II. and q, remain the same. Thus for $1 = II.,

-

By combining only the last two equations:

Also, since qel = q,, it follows that

and

It should also be noted that since the head coefficient II. and specific capacity q, have remained unchanged, the specific speed N,, is also the same. The efficiency would also be expected to remain the same, except for the influence of losses which vary according to laws other than the fundamental forms in the foregoing. Test results, such as those shown in Fig. 12, show the efficiency to be essentially constant. The principal results are again summariaed in the familiar rule that if the capacity changes in proportion to the speed, then the head changes in proportion to the square of the speed, the efficiency remaining relatively constant. Also, since the horsepower is equal to capacity times head, it is seen that the bralce horsepower varies directly as the cube of the speed, provided the efficiency is constant. The foregoing rules must be applied cautiously, taking into consideration deviations due to changes in Reynolds number, size effect, and cavitation. Also illustrated in Fig. 12 are curves representing the normal system head and the throttled discharge head. A particular condition of capacity and head is produced at the point of intersection of the pump performance

and finally

Usually the units of rpm and gpm are used. The symbol N,, is sometimes used for specific speed (to distinguish it from n,) and the expression becomes:

The specific speed as defined in the foregoing is not a purely dimensionless number, but this makes little difference since it can be made dimensionless simply by

~

.

~

(

I

I

I

I

I

I

I

I

I

I

~

~

FLOW RATIO= FLOW/ RATED FLOW fig. 13

Characteristic curves for a aeries of geometrically similar pumps

curve with the system head curve. It is customary to specify the rated pump performance at or near the best efficiency point (abbreviated BEP), and it w ill be noted that the pump illustrated in Fig. 12 has a BEP (capacity ra&o = 1.0) almost on the normal system head curve. The capacity-head curves of Fig. 12 tend to approach zero at relatively high rates of flow. It is uncommon to specify pumps for such a broad range of performance due to the deleterious effect of high velocities beyond the BEP, which may cause erosion of the internal parts, and the extreme low head which may cause an upset of the internal axial hydraulic balance. When applying centrifugal pumps to a specific set of performance requirements, an operating flow limit twenty percent beyond the BEP or rated point is commonly accepted. b. Performance Curves for a Series of Geometrically Similar Pumps. The same line of reasoning as described in the foregoing is used to predict the approximate performance of a new pump which is of a different sise but geometrically similar to one that has been tested. I n this case the advantage of plotting the capacity-head values.in ratio form is that such a dimensionless leadcapacity curve then represents approximately the performance of a whole series of geometrically similar pumps, and only the values corresponding to the design point need be calciiated from the similarity relationships. If &, H, N , and D are used as the test values at the designed capacity and speed of the original pump, then the corresponding values for any other similar pump of diameter Dl and speed NI may be determined as shown in the following equations: Q1 = Q

Nl Dl

(z)

and

& =H

(&)'

(26)

It sometimes is useful to plot the characteristics for both the original and the new pumps to the same scale. For example, a design might be laid down for a series of geometrically similar p,umps with a 10 percent range of diameters. Then, since the capacity usually can be varied from about 0.75 to 1.2 times the rated values with not over a 5 percent change in efficiency, this series of pumps could cover a fairly wide useful range of headcapacity performance without any .change in the,pbxnp speed. This is illustrated by the zone outlined with cross hatching in Fig. 13. Referring to the foregoing equations, it is seen that construction of the head-capacity curve for a geometrically similar pump of different size is obtained by varying the head as the square of the diameter and the capacity as the cube of the diameter, if the speedis held constant. This same figure also shows how a new useful range zone can be constructed from the original one by increasing the speed and using a different range of diameters. I n this way it is possible to constmct a series of slightly overlapping zones which would serve as a chart for selecting the required size and speed of pump for any desired combination of capacity and head. c. Factors Affecting Pump Efficiency. The d c i e n c y of various sises and types of pumps may, be plotted as a function of the specific speed N,, with the results shown in Fig. 14. The specific speeds were calculated on the basis of the installation conditions of service and are not necessarily the BEP for the actual pumps. Owing to such considerations,aa reliability, maintainability, first cost, space requirements, or choice of stock pump sires, pumps as selected are generally less efficient than those designed for optimum efficiency. Thus values of maximum efficiency are frequently found in practice that are

*

~

MARINE ENGINEERING LEGEND

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS I

I

I

I

I

I

41 3

I

A -RADIAL-FLOW CIRCULATING PUMPS 8- AXIAL 8 MIXED-FLOW CIRCULATING PUMPS C FEED PUMPS D- MAIN 8 AUXILIARY CONDENSATE PUMPS E CARGO W M P S

-

EFFECT OF VISCOSITY

-

I MAX. EFF. PUMPING WATER AT ROOM TEMP. 2- MAX. EFF. SAME PUMP WlTH OIL AT 6 0 0 SSU 3 MAX. EFF. SAME PUMP WlTH OIL AT 2 0 0 0 SSU

-

e 0.2

5 1x

I

a

u

0

0.2

0.4 RATIO=

0.6 0.8 1.0 ACTUAL FLOW RATED FLOW

1.2

1.4

Fig. IS Hfect of viscosity on single-stage, double-suction pump characteristics

SPECIFIC S@EED,N,

=

RPM fin

I

CEN

Fig. 14

MIXFLO

n'"

PROPELLER

Repmsentahve pump diciencies Venus specitlc speed for actual installations

five to ten points higher than those of Fig. 14, for pumps of low and medium capacities. The zones on Fig. 14 indicate roughly the regions to which the particular pump application is confined by experience and practice. The portion of zone A beyond 6000 specific speed for radial-qow circulating pumps represents a comparatively small number of installations where the puinps were selected for a capacity considerably in excess of the maximum efficiency point. Conversely, the range of zone B below about 6000 for mixedflow and axial-flow pumps represents a few cases where these types were selected for a capacity below the maximum efficiency point. In general the value of 6000 specific speed at maximum efficiency represents the dividing l i e between radial and mixed-flow pumps. Similarly, pumps above specific speeds of 7500 are of the axial-flow propeller type. The c w e s reveal that the efficiency rises to a maximum and then gradually drops as the specific speed is further increased. Also, up to a certain limit, higher

efficiencies are obtained with greater volume flows. Both of these effects are mainly due to the iduence of the frictional losses in the pump. . If the specific speed is maintained constant, then, as the capacity increases, the head loss due t b surface friction becomes a smaller percentage of the actual net head while the shock losses, clearance losses, and rotation losses remain a relatively constant percentage. In the case where the volume flow is maintained constant and the specific speed is varied, the percentage shock loss remains relatively constant while the percentages for clearance and windage losses decrease as the specific speed is raised. The percentage friction loss decreases rapidly at first, reaches a minimum, and then &dually becomes greater as the specific speed is further increased. When using the curves of Fig. 14 for multi-stage pumps, N,, must be computed for a single stage. When considering double-auction pumps, the practice of using one half of total capacity is not consistent and care must be taken to compare vyious pumps on the same basis. d. Viscous Liquids. The effect of viscosity on the characteristics of a singlestage doublbsuction pump is shown in Fig. 15. The head-capacity and efficiency versus capacity curves are plotted as ratios referred to the design conditions for pumping water. Two important effects of increased viscosity are seen. First, the head-capacity curve at a constant speed is lowered so that the speed would have to be increased for the pump to operate at the rated head and capacity. The efficiency also is greatly reduced, although the

I I

I ,

SUCTION LIFT

I "0

0.2

Fig. 16

u

I

I

0.4

I

I

I

I 1.2

I

1

I

I

+ DISCHARGE STATIC HEAD

I

0.6 0.0 1 .O FLOW RATIO = FLOW/:RATED FLOW

I C

I

i

1.4

Condensate pump characteristic curves showing effect of variable suction

epecific speed for a given capacity is higher because of Vl, (c) the impeller inlet peripheral velhity ul, and (d) the reduced head. Most importantly, the horsepower re- the inlet shock angle. For complete breakdown at zero quired is significantly increased, either overloading an shock angle (i.e., at about the designed capacity) the ,--relation is found to be existing driver or requiring a new larger driver. The values corresponding to the peak efficiency for V12 0.085 ulS each curve also have been plotted against specific speed Haw= 1.485 (27) 2g 2s on Fia. 14, where they follow a curve which is not greatly steeper than the effiEiency curves for constant ~ i ~ a c i t - Equation ~ (27) expresses the suction condition limits at low specific speeds. This resemblance supports the strictly as a function of impeller inlet design. For more foregoing conclusion that the reduced pump efficiencies general considerations it is found that the requirements associated with low specific speeds are due primarily to for similarity of impeller inlet conditions can be expressed greater frictional losses. Further information on viscous- by a dimensionless grouping faentical in form with the liquid performance may be found in references [9] and previously described 'specific speed. This parameter is [lo]. known as the suction specific speed S, and is obtained by e. Suction Lift and Cavitation. The maximum abso- replacing the total head H in the usual specific speed lute fluid velocity in the suction part of the system usu- formula by the net positive suction head H,. [ll]. Thus, ally occurs at the impeller inlet, hence this is the zone of in the customary units of rpm and gpm minimum absolute pressure. If at any point in this zone the vapor pressure of the fluid is reached, then a portion of the fluid will evaporate and form vapor pockets in the stream. These cavities disturb the flow stream and then I n a radial-flow pump where the suction and discharge collapse as they are carried into regions of higher pressure, thus producing noise, vibration, and rapid erosion zones of the impeller are clearly separated, the suction of the surrounding metal surfaces. This general behavior specific speed by itself is usually suf5cient to define the is known as cavitation and the necessity to avoid it im- cavitation limits independently of the discharge flow poses definite restrictions on the design and application conditions. However, when the inlet diameter apof centrifugal pumps. C~ndensatepumps are sometimes proaches the discharge diameter, the discharge flow condesigned to operate in the cavitation range as a simple ditions also have an influence on the cavitation condimeans of self regulation in spite of the many disad- tions, so that it becomes necessary to consider both S and vantages already mentioned. The effect of cavitation on the usual specific speed NEW.When considering doublethe pump characteristic curve is shown in Fig. 16. The suction pumps, care myst be taken to use one half of the beginning of cavitation is indicated at point B for ten- total flow rate. The curve's of Fig. 17-"~epr6sent %he upper limits of inch submergence and at point D for eightrinch subfor double-auction pumps handling mergence. Points C and E on the steeply falliig curves specific speed, NEW, clear water at 85 F at sea level, as published by the Hyare in the region of complete breakdown. Conditions for incipient cavitation and also for com- dradic Institute [9]. For other conditions of higher plete breakdown are expressed by correlating: (a) the temperature, for liquids other than water, and for singlenet positive suction head H , , defined as the total suction suction, mixed-flow, and d - f l o w pumps, reference [9] head in feet of liquid, absolute, less the vapor pressure of contains additional data. When considering the cavitathe liquid in feet, (b) the impeller absolute inlet velocity tion pe$ormance of a pump in relation to its application,

+

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

MARINE ENGINEERING

I

SPECIFIC SPEED. Nw

Ill1If

=

RPY M m Hv4

FOR DOUBLE SUCTION PUMPS

Fig. 17 Upper limita of speciflc speeds for double-suction pumps handling clear water at 85 F at sea level

the user must be careful to distinguish between the net positive suction head available (abbreviated NPSHA) and the positive suction head required by ,the pump (abbreviated. NPSHR). For satisfactory performance, NPSHA must exceed NPSHR by a certain margin which depends on the particular application. An appreciation of the extent of analysis and testing that is required in pursuing the solution to cavitation problems may be obtained from references [12-141. '

1.3 Pump Characteristics a. D e s i g n F e a t u r e s . The physical components of a

centrifugal pump consist of one or mor&+pellers, casing, shaft, bearings, stuffing boxes, coupling, minor components associated with the rotating or stationary parts, instrumentation, controls, and supporting auxiliaries. Energy is imparted to the pumped fluid by the impeller, which may be of any of the types described in Section 1.1. All other components play a supporting role in the makeup of a particular pump design. The casing guides the fluid from the suction pipe to the eye of the impeller, collects the fluid exiting from the impeller outlet, guides the fluid to the discharge pipe, and acts as the major pressure boundary in the manner of a pressure vessel. Casing designs, classified as to type of flow, are described in Section 1.1. In addition, casings may be classified as to the type of collector, i.e., volute casings which have the form of a volute or spiral in the direction of flow from the impeller, circular casings which have a constant cross section concentric with the impeller, or diffuser casings which have a multiplicity of vanes to guide the fluid exiting from the impeller and convert the developed velocity head into pressure. .Shafts of centrifugal pumps are designed to transmit the torque needed fox &king the impeller,.to resist the

bending loads that occur due to internal hydraulic forces, external alignment forces, and weight of parts, and to control critical shaft vibratory conditions. Practically all pumps operating at 3500 rpm or less are designed with a first critical speed above the running speed. For speeds above 3500 rpm, such as for boiler feed pumps, the operating speed may be above the first critical speed but suitr ably below the ~econdcrikical speed. Bearings are of the antifriction type (ball or roller bearings) or of the sleeve type. (hydrodynamic type). Ball bearings, such as depicted in Fig. 2, find frequent application. Pumps running above 3500 rpm usually use journal bearings of the sleeve type and Kingsbury thrust bearings, as depicted in Figs. 4 and 5. Figure 3 shows a combination of types, that is, a ball bearing for the thrust and upper journal bearing, and a sleeve bearing at the lower or internal journal. In the latter case, the ball bearing is grease lubricated and the internal sleeve bearing is water lubricated. The ball bearings shown in Figs. 2, 3, and 8 are lubricated by a self-contained system; that is, fittings are provided for injecting fresh lubricant and draining used or excess lubricant. Seals are provided to retain the grease or oil within the housing. Sometimes cooling water is circulated in a surrounding cored pas-sage, as in Fig. 2. The bearings for the pumps in Figs. 4 and 5 are lubricated by a separate pressurized system. The pump in Fig. 7 contains an internal water-lubricated journal bearing immediately above the impeller. However, it has no thrust bearing, the axial hydraulic t h r u ~ t and rotor weight being carried by the thrust bearing in the driver. Water-lubricated bearings are satisfactory only for clean-water service. Packed stuffing boxes prevail in marine centrifugal pumps, although mechanical seals and packless boxes are finding wider usage. The s t a n g boxes for the pump in

'

i

I

Fig. 2 are subjected to positive suction pressure and therefore no special provision is needed to ensure a steady trickle of leakage through the packing rings. When the atuffing box could be subject to lift or vacuum conditions, such as with the pump in Fig. 3, provision must be mqde to inject sealing water under positive pressure at a point ~pproximately midway between several rings. Some sealing water \vill therefore enter the pump, and a small portion nil1 trickle from the box. The packed stuffing boxes in Fig. 5 require water cooling jackets due to the high temperature of the boiler feedwater being pumped (200-350 F), and due to the greater friction heat generated by the higher speeds. The pump in Fig. 4 contains packless boxes. Condensate is injected into closerunning-clearance serrated bushings at both ends of the pump. A small portion of the injection water enters the pump, but the major part leaks out the ends, collects in the drain chambers, and passes out the large drain connections. The pump depicted in Fig. 8 contains a mechanical seal. Many commerical pump designs can be equipped with either conventional packing, mechanical seals, or packless boxes. Couplings connect the pump to its driver. Commonly used types are pin and buffer (Figs. 2 and 3), gear (Fig. 4), rigid or solid (Fig. 7), and hub and spider (Fig. 8). Couplings are flexible to accommodate small misalignments between the pump and driver, except for the single-bearing pump of Fig. 7 which requires a solid (nonflexible) connection. Couplings for small pump shafts and drivers usually have a straight pressed or shrink fit (Fig. 8). Medium size and larger pumps usually have the coupling hubs mounted on a tapered fit (Figs. 2 through 7) for ease of removal. Marine centrifugal pumps are almost universally fitted with packing sleeves (Figs. 2 through 4, and 8) and journal sleeves in water-lubricated bearings (Fig. 3). A small flinger devise is usually fitted on the shaft immediately outside the stuffing box for throwing off water migrating along the shdft (Figs. 3, 4, 5, and 8). Casing wearing rings are universally installed in all types, whereas impeller rings are generally furnished only on medium and larger size pumps up to 3500 rpm (Figs. 2 and 3). High-speed boiler feed pumps (Figs. 4 and 5) generally do not have impeller rings, due to the relatively higher stresses in the impeller, and the difficulty of maintaining a reliable fastening 'at the joint. Centrifugal pumps are usually fitted for a minimum of instrumetation consisting of discharge and suction pressure gages and a lubricant supply indicator. Additional instrumentation may consist of a speed indicator, vibration indicator, pressure gages for seal water, intermediate-stage pressures, leakoffs, and lubrication systems, thermometers or thermocouples for indicating temperatures of the pump fluid, lubricant, bearings, and casing, and' flowmeters to indicate pump flow, recirculation flows, or leakofl flows. Control devices are usually associated with the system (i.e., recirculation control, pressure or temperature control of injection water for seals, remote-operated suction

41 5

and discharge valves, low suction pressure trip) but may be associated with the diiver (i.e., automatic or remote s t a ~arid t stop or shutdown upon loss of lubricant). The wide range of uses and characteristics of centrifugal pumps on shipboard are described in the following paragraphs. b. C o n d e n s e r C i r c u l a t i n g P u m p s . Both the main and auxiliary condensers re@ire a largc+volume flow of circulating water at relatively low heads; therefore, the specific speeds of. pumps for this purpose are high. For the same con9tions of capacity and head, the higher the specific speed, the lower the size and weight of the pump. Thus, circulating-pump designs of the radial-flow type were developed in the direction of increased specific speed until the limit due to cavitation was reached-at about N,, = 6000 at maximum efficimcies for doublesuction pumps. At the same time, mixed- and axial-flow main circulatr ing pumps are used widely for high-speed'qhips fitted with scoop circulation. In such cases the circulating pumps are used only for very low ship speeds and for maneuvering. Since they are operated at full capy5ty for only short periods, the specific speeds are increased beyond the usual cavitation limits. These pumps have slightly lower efficiencies due to the higher specific speeds, but this is unimportant because of their infrequent oper* tion, and is fully justified by the large savings in weight and space. Motor-driven axial- and mixed-flow pumps require a larger motor than indicated by the design conditions, because the steeply rising head'yaracteristic results in a rising horsepower curve as the capacity is decreased. An error on the low side in estimating the system head requirements would result in overloading the motor, if its ratihg were very close to the pump brake horsepower at the design capacity and head. Radial-flow circulating pumps are available with a fairly flat, non-overloading horsepower characteristic, and these units can be used safely with a motor which is rated close to the pump horsepower. A typical circulating pump of the vertical double-suction type, for smaller vessels, is depicted in Fig. 2. Circulating pumps of the vertical mixed and axial-flow types for larger vessels are depicted in Figs. 6 and 7. Seawater circulating pumps for submarines are of the radial-, mixed-, or axial-flow types, but differ in construction $ that their casings must be designed to withstand the ambient pressure due to submergence, shaft seals must be suitable for the ambient pressure, and thrust bearings must be sized for possible high axial thrusts. Circulating p u q ~ ~ r e e c ~ m m ofound n l y to be constant-speed and motor-dr~venin c;mme~cial vessels, although multi-speed motors are sometimes used. Steam turbine drives are often used on naval vessels due to the flexibility of the variabIe-speed driver and the reliability of steam as a power source. Additional details regarding circulating applications may be found in ~eferende r151. . c. C o n d e n s a t e and rain P u m p s . Continuing deA

41 6

MARINE ENGINEERING

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

where p, = atmospheric pressure, psia pop= vapor pressure, psia p = specific weight of liquid, pcf ha = total suction head, ft

SINGLE SUCTION PUMP CAPACITY, GPM

Fig. 18

Capacityspeed limitations for condensate pumps wilh h a f h *rough eye of impeller

velopments in marine power plants have increased the condensate pump head requirements. A typical condensate system consists of an air ejector and gland vapor condenser and a deaerating-type feedwater heater having a shell pressure of 10 to 60 psig. Such installations, in general, require a two-stage condensate pump. The condensate pump (sometimes referred to as the condenser or hotwell pump) is required to develop a total head made up of the sum of: the friction loss in the system due to heat exchangers, piping, fittings, and valves; the difference in elevation head between liquid levels in the hotwell and receiver (usually a deaerating-type feedwater heater); and the difference in pressure levels between the hotwell (usually being under vacuum) and the receiver which is usually at a shell pressure of 10 to 60 psig. Because of the moderate to small capacity of these units, the impellers are generally of the sin&e+uction type. In vertical pumps the submergence at the top inlet eye becomes too low if a double-suction construction is used. With two-stage pumps both suctions are placed on the inner side of the impellers, that is, the impellers are positioned "eye to eye," so that the shaft passes through the casing at the upper part and the single stufFing box is usually under discharge pressure. A twostage vertical condensate pump is depicted in Fig. 3. The specific speed is inherently low because of the low capacity-head ratio, and is limited by the available motor speed, which is 3500 rpm for a-c motors. However, only condensate and drain pumps of relatively low capacity (generally below 100 gprn), such as used with turbogenerator condensers and distilling plants, can operate

at this speed due to the low values of net positive suction head available. The demand for condensate pumps to operate at reduced submergence requires lower speeds, so that the majority of the condensate pumps operate at 1750 rpm or less. Consequently, quite low efficiencies are to be expected with condensate pumps; this is shown by zone D of Fig. 14. Limitations of capacity and speed for condensate pumps are given by Fig. 18 for single-suction pumps; for double-auction pumps, the capacities read from Fig. 18 should be doubled [9]. For the small capacity of 100 gpm for the 3500 rpm pump mentioned earlier, the suggested NPSH is four feet. Computing the suction specific speed, as defined previously in this section, S is found to equal 12,360. For a condensate pump of 330 gpm operating with an NPSH of 2 feet, the suggested speed is 1150 rpm and the computed S is 12,400. In general, condensate pumps perform with a range of suction specific speeds of 12,000 to 18,000, whereas pumps for other applications p e r f o p with suction specific speeds below 12,000. In applying condensate pumps, a distinction must be made between the terms submergence, NPSH, and suction head. Submergence relates the liquid level in the hotwell to the elevation setting of the pump. It is a static dimension, and it is customary to measure it to the centerline of the suction nozzle of vertical pumps, and to the shaft centerline of horizontal pumps. Suction head refers to pressure above atmospheric. NPSH is the total suction pressure above the vapor pressure of the liquid on the absolute scale [9]. Thus, denoting NPSH by Ha,, -

When the pump takes suction from a hotwell where the prevailing pressure equals the vapor pressure corresponding to its temperature, the NPSH is the difference in elevation between the liquid level and the datum (suction nozzle), minus the entrance and friction losses in the suction piping. Actually, condensate pumps operate a large part of the time at values of NPSH below the design value. This is because condensate pumps are generally driven by constantrspeed motors and there is no external response to a change in the amount of steam condensed. Turbinedriven units are controlled only by a speed-limiting governor so that their behavior is similar. The results of a reduction in the amount of steam condensed is shown in Fig. 16. Point A is the maximum-load operating condition corresponding approximately to the pump rating. This point is at the intersection of the pump headcapacity m e with the system m e . Assume there is a sudden reduction in condensation rate to the value E. At first the pump continues to deliver at capacity A, thus reducing the hotwell level and submergence. When the submergence has lowered to the value corresponding to the cavitation curve BC, the pump will be delivering at a capacity corresponding to point C, and finally equilibrium will be reached at the submergence and capacity corresponding to point E. A condensate pump operating on the vertical head curves, such as BC or DE, is said to be "operating in the cavitation break." Thus the part-load operating conditions for condensate pumps are severe from the standpoint of cavitation, and special consideration is given to this in the impeller and overall design. Aside from the possibility of rapid wear, these conditions are also difficult from an operating standpoint. If the cavitation part of the pump curve is very steep, operation is likely to be unstable with large fluctuations of the pump discharge pressure. There is also the possibility that the pump will become completely vapor bound and deliver much less then the required capacity until the submergence has risen considerably above the normal value, thus resulting in intermittent or slug flow in the condensate piping. An alternative method of operation employs an automatic level control. This arrangement causes part of the pump discharge flow to be bypassed back to the hotwell, thereby maintaining a prescribed level of condensate in the,hotwell and ensuring an adequate NPSH at the pump. Pump operating conditions sometimes are aggravated by insistence upon excessive margins in the specified head and capacity of the pumps. Better operation and longer life will result if the capacity and pressure loss require-

417

ments are carefully selecked and the design maxgins are held to moderate values. 'In this connection it should be noted that great care is necessary in the installation to prevent air leaks, as tests show conclusively that very small leaks will cut down the capacity by a marked d e pee. For the same reaaon it is quite important to avoid forming air pockets in the syction line. A vent connection located at or near the suction nozzle and connected to the condenser assists in cleaxing vapor' from the im,.' peller eye. d. Boiler Peed Pumps. The etliciency characteristics of boiler feed pumps are shown in Fig. 14. It is seen that the efficiency zone is slightly higher than for condensate pumps because of the somewhat larger capacity required. For the same capacity and specific speed, aciencies of the two are about the same despite the larger number of stages in feed pumps. The efficiency is about the same for volute and diffuser-type pu,mps, the choice of type being principally a matter of the pump manufacturer's practice or the customer's preference. Advances in design have increased pump speeds and consequently the maximum head per stag6 so that twostage and single4tage feed pumps are available for boiler pressures up to 1200 psig. Centrifugal boiler feed pumps are not widely used for capacities below about 100 gpm because the combination of low capacity and high relative velocity (due to the high head per stage) results in impeller passages which are quite small and therefore are more likely to become clogged or restricted by smqll-particles or boiler water deposits. Figure 4 depicts a horizontal two-stage boiler feed pump and Fig. 5 depicts a horizontal four-stage pump. Single-suction impellers of boiler feed pumps develop a large axial thrust due to the pressure diffeience on the two sides of the impeller unless both back and front wearing rings are fitted. For reasons of reliability and maintainability, however, boiler feed pumps do not have back wearing rings, nor do they have impeller rings at the eye. Where an even number of impellers is used, these axial forces may be balanced by arraaging the impellers to face each other. For more than two stages, this arrangement requires some of the interstage passages to cross over other stages, thus complicating the casing design considerably. A balance drum with labyrinthdthrottling paasages is often used to compensate for the axial thrust of several impellers, but this arrangement will not always balance the thrust under all operating conditions; therefore, an external thrust bearing is also required. One means of automatically balancing the axial thrust unde?5al~~condit,ns of_-operationis the balance disk, which varies the clearance in the throttling element and hence the balancing pressure by permitting a slight axial movement of the pump shaft. Operation of boiler feed pumps under cavitating conditions is never permissible because of the risk that the pump suction might become vapor bound. If this happens, all the water in the pump may be evaporated quickly owing to the large energy losses in the pump

*

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

MARINE ENGINEERING

relative to the volume of water it contains, and serious damage to the running parts probably will result. For this reason most feed pumps are provided with low-suction pressure trips. For the same reason boiler feed pumps cannot be operated safely at conditions approaching shut-off and are provided with a recirculating line which discharges from 5 to 15 percent of the rated capacity back to the deaerating feed tank. The chances of complete vapor binding at rates approaching shut-off are greater than in condensate pumps, not only because of the larger energy losses in the pump, but also because a t higher water temperatures a small rise in temperature results in a greater increase of the vapor pressure. The limiting suction conditions must depend, therefore, not only upon the specific speed and suction specific speed, but also upon the temperature of the water, owing to the greater rate of vapor pressure rise. Figure 19, as presented in reference [9], depicts the recommended minimum NPSH for pumps handling water at temperatures of 212 F and above. Since, for boiler feed pumps, the specific gravity is usually less than 1.0, care must be taken to include this factor when converting actual suction pressures in psi to values of NPSH, and when converting suction and discharge pressures to total head. A boiler feed pump is required to supply water a t a pressure sufficient to overcome the sum of the highest safety. valve setting of the boiler, the pressure due to the static elevation of the boiler above the pump, and the friction head loss in the system of piping, valves, and fittings. It is usual practice to plot the system head curve and superimpose on it the head-capacity curve of the pump as depicted in Fig. 20(a). A turbinedriven boiler feed pump may employ any of the following three types of governors: 1. Constant-speed governor only. 2. Constant-pressure governor to maintain a constant discharge pressure at all capacities. 3. Differential-pressure governor to maintain a specified differential-pressure across the feedwater regulator. These governors are described further by the simplified diagrams of Fig. 20. In the constant-speed governor system, Fig. 20(b), the feedwater regulator throttles all the difference between pump discharge pressure and the required pressure. The throttled pressure differential represents a large waste of power, hence a constant-speed governor is rarely used alone; it is usually provided along with a constant-pressure or differential-pressure governor, in which caqe the speed governor provides a secondary means of control when the pressure governor may be temporarily out of service. In the constant-pressure governor system, Fig. 20fc), the governor controls the turbine and pump speed to provide a specified constant pressure at the pump discharge. The feedwater regulator throttles the variable difference between the constant pressure and the system head curve. In the differential-pressuregovernor system,-Fig. 20(d), the

SYSTEM MAD CURVE WlTH PEEDWATER REGULATOR FULLY

H-o CHARACTERISTIC OF PUMP

500 400

P OPEN

OESIQN -HEAD

300 200

-----

~ O I L E RPRESSURE

I

L C

a o

+.I00

W

r 90

I

I OESIGN

80

0

g

* 5 In

PLUS STATIC ELCVATION

70 60 50

CAMCITY

CAPACITY

(a1 TYPICAL BOILER FEED SYSTEM

40 30

W

2 k

20

SPEED

V)

B

k

I0 I

CAPACITY, GPM Fig. 19

Net positive suction head for single-suction, centrifugal, hot water Pump'

governor oontrols the turbine and pump speed to provide a specified differential pressure across the feedwater regulator. The pump discharge pressure coincides with the system head curve. , The diagram in Fig. 20(e) shows the system head curve for the boiler feed system of a nuclear steam plant. In this case the pump is called upon to operate at its highest speed at the low-flow condition. The operation of the pump is stable, however, since, even though there may be an inflection in its speed curve, there is no inflection in its horsepower curve throughout the range of needed capaoity. When applying boiler feed pumps, the method and amount of recirculation must be coordinated with the method of pressure governing. A widely accepted practice [5] is to hold the temperature rise at low flows to 15 deg F, though rises of 20 or 25 deg I? may be found in use. A boiler feed pump operating under the control of a differential-pressure governor system requires a certain amount of horsepower at the minimum-flow condition, aa indicated in Fig. 20(d). If for some reason the control system is made inoperative and the pump then operates at constant speed, the horsepower to be dissipated is a larger amount, as shown in Fig. 20(b). Thus the reek culation flow must be sized for the \\?orst condition that the system may encounter. In addition, consideration must be given to whether the recirculation flow is permitted only a t or near the minimum-flow condition, or whether it is continuous. The various types of recircualtion devices and systema may be summarized as follows. All systems require a pressure breakdown device which may be a singb plate orXce, a multiple orifice (spool type), or a small-bore tube (friction tubing). The pressure breakdown device

SYSTEM HEAD CURVE wlm FEEDWER RESULATQR

/:

1

---ELEVATION

I'

URE

-*---

BRAKE HORSEPOWER

IMSIGN

CAPACITY

CAPACITY

tb) CONSTANT-SPEED GOVERNOR SYSTEM

H-Q GHARACTERISTIC OF PUMP-CWSTAMT SPEED

WlTH F E E D W E R

is installed in a line connected from the pump discharge back to the deaerating feed tank. For continuous recirculation, no other devices are fitted, except stop valves for isolating the line. For automatic intermittent recirculation, a control device is required. This may be a diaphragm-operated valve, controlled by an air signal from a flow transmitter, which in turn is connected to a flow-measuring orifice in thwpump discharge line. Another type is a special discharge check valve in which the movement of the-internal check operates a pilot valve that opens or $lcises the recirculation line. Most marine boiler feed pumps are driven by steam turbines. Whereas Figs. 4 and 5 depict coupled designs, single- and two-stage types that have pump shafts common with the turbine shaft are frequently found. Smallcapacity pumps for waste-heat boilers or small auxiliary boilers are usually of the end-suction type, Fig. 8, and are motor driven. Some ir~stallationsutilize vertical multi-stage pumps, either motor or turtine driven, and some vessels have the feed pump driven by the main propulsion turbines or by the turbogenerator set. e. Feed Booster Pumps. Feed boosterr-pumps - are needed in connection with deaerating heater type feed systems where deaerating heater space at arestrictions sufficient static prevent elevation locating above the the boiler feed pump suction to provide the required net positive suction head. In the absence of a large static submergence a t the suction, the required cavitation limits are satisfied by the booster pump discharge pressure. With this arrangement a relhxivelv low NPSH condition occurs at thevfeed booster However, these pumps usually are designed for much lower speeds and heads so that the impeller is larger and the suction velocities are lower than in boiler feed pumps. The result is that, for about the same suction specific speed, 8,

PLUS STATIC

I

-

I

CAPACITY

le 1 CONSTANT-PRESSURE GOVERNOR SYSTEM

H -Q CURVES AT SPEEDS ABOVE 100%

H-Q C M M E R I S T I C OF PUMP-160% s m

H-Q CHARACTERISTLC 0 W

REWLAlOR ,

r

V A R I a I L E PRESSU# VARIABLE SPEED 5 STATIC

t

I

I CAPACITY

(dl DIFFERENTIAL-PRESSURE GOVERNOR SYSTEM Fig. 20

CAPACITY

(el NUCLEAR STEAM POWER PLANT

Boiler feed pump and system characteristic curves

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

MARINE ENGINEERING

a lower net positive suction head is permitted for the booster pump. Also, the possibility of complete vaporbation is reduced in the booster pump because the ratio of energy loss to pump volume is so much smaller than for boiler,feed pumps. Feed booster pumps are in many respects similar to condensate pumps. Booster pumps do not operate in the cavitation break, a s mentioned previously for condensate pumps, 'as a positive stable performance is required in order to provide the NPSHR for the boiler feed pump. The capacity of the booster pump is essentially controlled by the boiler feed pump, which in turn is controlled by the boiler feedwater regulator. A booster pump must be provided with a recirculating line, usually piped back to the deaerating feed tank, to protect the booster pump when the feed pump is not in operation. A typical twostage booster pump is similar to the condensate pump depicted in Fig. 3. A singlestage booster pump would be similar to that of Fig. 3 except that there would be no second (upper) stage. Booster pumps frequently use only ball bearings, due to the severe duty imposed upon them by tbe higher water temperatures and danger of flashing w h i ~ bmake water-lubricated bearings less suitr able. f. Cargo Pumps. Representative efficiencies of cargo pumps are shown in Fig. 14. Axially split casing types of pumps, with double-suction impellers, similar to that depicted in Fig. 2, are used in the capacity range of 2000 to 20,000 gpm at total heads of 500 ft. This type of pump can be horisontal or vertical, being driven by a geared steam turbine in the main machinery space, with the driving shaft extending through a bulkhead or deck stuffing box to the pump located in an adjacent pump room; or the pump may be of the deepwell type. The geared turbine drive permits the use of a highly efficient turbine far the large power requirements and permits compatible ,matching of the most desirable pump and driver speeds. It provides variable-speed operation in order to obtain high speeds for producing the highest heads or lower speeds to suit lower values of NPSHA, thus enabling the pump to respond to the wide range of system head characteristics that are encountered under different operating conditions. The horiaontal-shaft arrangement has the disadvantage that misalignment may ensue under different conditions of vessel loading. To overcome this, and to reduce fore-and-aft space requirements, vertical doublesuction pumps are used in many vessels. Here the pumps are driven by a turbine and reduction gear located on a machinery flat extending from the main machinery space, directly over the pump room. Vertical shafting is used to connect the pumps and drivers. Cargo pumps are required to pump a variety of volatile liquids and thus require means to ensure continuous pumping under possible vapor-binding conditions. Most older vessels used reciprocating steam pumps, which were operated during the stripping operation. This type of pump, being positive displacement,was capable of pumping vapor6 as well as liquids. It wm operated coin-

1

I

1

i LEGEND 1 2 3 4 6

8 7 8 B 10 11 12 13 14 15 18

SUCTION BELL FIR-AGE IMPELLER (AXIAL.FLOWTYPEI FIRSTSTAGE DISCHARGE BOWL EWALlZER LINE SPACER PIECE SECONDSTAGE SUCTION BELL SECOND STAGE IMPELLER (MIXED-FLOWTYPE) SECONDSTAGE DISCHARGE BOWL THIRD,STAGE SUCTION BELL THIRD STAGE IMPELLER (MIXED-FLOWTYPE) THIRDSTAGE DISCHARGE BOWL VENTLINE SUPPORT COLUMN WAFT PACKING BOX DISCHARGE HEAD 81DRIVER SUPPORT

Fig. 21

Deepwell-type centrifugal pump

cidently with the centrifugal cargo pump, where it effectively removed the vapor collection at the suction of the centrifugal pump. A variety of systems for removing vapor is found in modern vessels. One such type employs an integral priming or scavenging impeller located at both ends of the regular impeller. The vapor discharged from the priming impellers is piped to a separator tank, from which any liquid present may drain back to the pump suction, and from which vapor is piped to the overboard vent system. Another type makes use of a separate vacuum pump which pumps vapor from a separator tank located in the main pump suction line (see

,

Fig. 21 of Chapter 18). The vapor passes through an interceptor tank, where any liquid carryover is removed and drains back to the main pump suction. The discharge of the vacuum pump is piped to the overboard vent system. Associated instrumentation and con.trols cause the vacuum pump to start and stop as the amount of vapor accumulates in the separator tank. Another type, which is used with the deepwell type of pump, incorporates an automatic priming valve. In addition to steam turbine drives, cargo pumps may be driven by electric or hydraulic motors, by diesel engines, or by gas turbines. A common arrangement is a vertical deepwell-type pump driven by a vertical explosion-proof votor, or by a horizontal diesel or steam turbine driviqg through a rightrangle gear. Such a design, depicted in Fig. 21, is of the single-suction type and is multi-stage. It may contain a specially designed firstr stage impeller to suit the particular suction conditions. For the handling of low-temperature liquefied gases, cargo pumps are almost without exception of the submerged deepwell type, or of the completely submerged canned-motor type. The deepwell type is mounted vertically, with its driving motor located on the deck above. It is also used for chemical cargoes. The canned-motor type is mounted vertically with the motor a t the bottom of the cargo tanks. The motor windings and bearings are completely submerged in the pump fluid. Both types usually contain a special design of suction impeller compatible with the low values of NPSHA. Additional discussi~nregarding cargo pumps is given in references [lo], [16], [17], and [18]. g. Bilge and Ballast Pumps. The general requirements of bilge pumps are similar to those of condensate pumps in that, due to the liftrequired of bilge pumps, the suction pressure is considerably below atmospheric. In comparison, the capacity is larger and the discharge pressure is less. The impeller passages must be much larger in order to pass dirt and bilge debris. As no static submergence is available to fill the pump with water, an auxiliary priming devide is required. This usually ia a positive-displacement air pump either directly attached to the main pump shaft or operated separately. An air float valve is used to isolate the air pump after priming has been accomplished, thus avoiding the chqrning loss of this element during normal operation. Some vessels employ an independent central priming pump for all bilge and ballast services. Bilge and ballast pumps are usually of the axially split type, either horizontal or vertical, similar to the pump depicted in Fig. 2. h. General Service. Centrifugal pumps are used for many minor and supporting services for capacities up to 5000 gpm and for total heads up to about 350 ft. Typical applications are for fire main, flushing, cargo oil tank cleaning, refrigeration condenser cooling, distilling plant feed, condensate, brine overboard, fresh water, and sanitary systems. Pump construction for these services is

42 1

usually that of the axially split casing type, either horizontal or vertical, similar to that depicted in Fig. 2, or of the end-suction type as depicted in Fig. 8. i. Primary Cooling Service. Pressurized water reactor systems require a circulating pump that is of the endsuction type with the impeller mounted directly on the motor shaft. The bearings are lubricated by the pump fluid, and the motor'windkgs are isolated in a separate casing. j. Jet Propulsion. Pumps are finding increased usage as pro dlsion devices, either as the main propulsion device or asPan auxiliary device such as a bow thruster, These pumps are usually of the mixed-flow or axial-flow type. The particular design depends on the exact matching of the pump capacity, total head, speed, and type of driver, which may be an electric motor, internal combustion engine, or gas turbine. Additional discussion of jet propulsion applications of pumps is contained in refer, ences [I91 to [21]. \ k. Materials. Centrifugal pumps for merchant vessels usually are made entirely of nonferrous materials when used for seawater applications. Cashgs and impellers are generally of cast and machined bronze. Shafts and sleeves are generally of nickel-copper alloys, suitably hardened. Sleeves and wearing rings are of bronze or nickel-copper alloys, depending on the severity of service conditions. Occasionally impellers and the internal surfaces of casings are coated with one of several commercial coatings available. Seawater casings subject to submergence pressure aboard submarines are cast of highstrength alloy bronze or of a.8opper-nickel alloy. For freshwater services, cast iron is occasionally used for casings, but bronze is the most common casing material for condensate and other low-pressure applications. Various grades of stainless steel are used for high-pressure boiler feed pump casings. Impellers for low-pressure services are of bronze or nickel-copper alloy and are of stainless steel for boiler feed service. Shafts for lowpressure services are of carbon steel or stainless steel and are of stainless steel for boiler feed service. Sleeves and wearing rings are of bronze, nickel-copper alloys, or stainless steel depending on the severity of the service. Materials for cargo oil pumps must be suitable for seawater, as they are often used for ballasting service. Pumps for low-temperature liquefied gases are usually of stainless steel, the particular grade depending on the low temperature encountered. Wearing rings for the caaing and impeller are always made of dissimilar grades of bronze or other alloys and are of different hardnesses to provide the best wearing properties at the close running cleyances. The same practice is followed for th%rotatiig Eidd s t a h n a r y parts of pressure breakdown labyrinths and balancing disks and drums. Bolting is usually made of bronze or a nickel-copper alloy for low-pressure services and of low-alloy steel for boiler feed pumps.

422

MARINE ENGINEERING

Sectien 9 Reciprecating Steam P~mps 2.1 Ckrssificdion and Types. The use of recipmcatr ing pumps in the marine field has diminished such that they are now primarily used only for specialized purposes. Reciprocating pumps are of the positivedisplacement type; that is, pumps which displace a constant volume of fluid from the suction to discharge port, for each stroke or revolution of the driven shaft. Reciprocating pumps are basically of two types: ( 1 ) the steam-driven directr acting type, where the steam-end and the reciprocating pump are built together as a unit with the motivating Power being provided by the steam-end; and ( 2 ) the Power Pump t.We where the reciprocating action of the Pump is provided by the rotary motion of an external prime mover and is converted to reciprocating motion by crankshafts, eccentrics, and cam plates, piston rods, etc(see Section 3 ) . Regardless of the type, the Pump end is classified the same. A reciprocating steam pump consisting of single steam and liquid cylinders and pistons with one piston rod is known as a single or simplex pump. Figure 22 shows a vertical simplex double-acting type. Where two such pumps of identical size are arranged side by side and the steam valve of one side obtains its motion from the piston rod of the other side, the unit is known as a duplex pump. Figure 23 depicts a vertical duplex double-acting type of pump. Both of the pumps shown in Figs. 22 and 23 are known as vertical pumps because the axial centerline of the cylinders is in a vertical position. Horizontal types are those having the axial centerline of the cylirlders in a horizontal position. The size of a reciprocating steam pump is described by giving first the steam cylinder diameter, then the water cylinder diameter, and finally the pump stroke, all in inches. 2.2 Flow Rate and Piston Speeds. A reciprocating steam pump produces a constant rate of flow throughout most of each stroke; however, the flow is reduced abruptly at the end of each stroke as the piston decelerates and reverses direction. Figure 24 shows, in relative terms, the maximum and mean flows for a simplex pump and duplex pump. It is desirable to fit air chambers a t the discharge to smooth out the flow pulsations to acceptable values. Basic piston speeds, as recommended by the Hydraulic Institute [9], are shorn in Fig. 25. The curve applies to simplex and duplex pumps of conventional design operating on cald water or on liquids possessing a viscosity of 250 SSU or lees. However, for handling viscous liquids, and when pumping hot water, lower piston speeds must be used; for recommended values, consult reference [Q]. Where U is the piston speed in fpm, Q is the volume rate of flow in cfs, D is the liquid piston diameter in feet, and S is the slip ratio, the piston speeds may be calculated by one of the following formulas:

For simplex pumps:

u=-

WQ

(30)

(*/4)D2(1 - S)

pumps:

U

=

'

-

*WQ

(31)

(u/4)D2(1- S )

2.3 ~ ~ l steom ~ + i pump ~ ~ cylinder sizes, The steam piston diameter must be proportioned torrectly in relation to the liquid piston diameter to obtain the desired pump pressure for a given steam pressure. Since the steam cylinder can have no cutoff, due to the negligible inertia of the moving parts, the pressure at the end of the stroke is less than the steam line pressure only because of friction and shock losses. The required ratio of steam piston area to liquid piston area can be calculated from the following expression: Da2 Dm2 &(pa

PW

- p,

--

(32) ~

1

)

where

p, = net liquid pressure, psi p, = steam inlet pressure, psia p, = steam exhaust pressure, psia pl = steam pressure loss entering and leaving steam cylinder, psi D , = steam piston diameter, f t D , = liquid piston diameter, f t Em = mechanical efficiency I n selecting cylinder diameter ratios, it is common practice to add a margin of 10 to 15 percent to the values obtained by calculation. 2.4 Mechanical and Volumetric Efficiency. Average values for the mechanical efficiency of reciprocating steam pumps are shown in Fig. 26. The volumetric efficiency is given as a loss in percentage of the displacement and is then called "slip." The slip a t rated piston speed for new pumps averages from 2 to 5 percent, but this value increases as the pumps wear in service, so that far design purposes it is better to allow for about 10 percent slip. The slip values at different rated pressures are based on pumps with packings designed for these pressures. 2.5 Steam Consumption. The approximate steam consumptio~lof a reciprocating pump at any given rate of fluid flow and total pressure is readily calculated from a consideration of the equilibrium of steam and liquid forces on their respective pistons at the ends of the discharge stroke. At this instant these forces, with due

a

1 2 3 4 6 6 7 8 0

Fig. 22

LEGEND

STEAMCYLINDER STEAM CYLINDER l STEAM ROD LIQUID CYLINDER - LIQUID CYLINDER LINER LIQUID PISTON LIQUID PISTON ROD VALVE SERVICE LIQUID PISTON P ~ K I N G

Vertical simplex pump

allowance for mechanical friction, are equal and the steam cylinder terminal pressure may be determined accordingly. The cubic content of the steam cylinder divided by the specific volume of steam in the cylinder gives the weight of steam contained in the cylinder at this inetant, and this weight of steam multiplied by the double strokes per minute of the pump gives the approximate steam consumption for each end of a double-acting cylinder.

LEGEND STEAM CYLINDER STEAM ROD LIQUID CYLINDER LIQUID CYLINDER LINER LIQUID PISTON ROD LIQUID PISTON LIQUID PISTON PACKING LIQUID VALVE SERVICE AIR CHAMBER

Fig. 23

Vertical duplex pump

The actual specific v o l p e of-the &earn a t the terminal pressure will depend upon the amount of-wndensation occurring during the stroke, and the steam leakage during the stroke also must be considered. It is customary to assume that the steam is in a saturated condition a t the terminal pressure and determine a corresponding overall condensation and leakage factor from actual steam consumption tests. On this basis the steam consumption may be obtained as follows. Using the nota-

MARINE ENGINEERING

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

PUMP STROKE Fig. 26

~ 1 24 ~ .

~

1

the following additional notation

w = pump steam consumption, lb/hr Q = volume rate of liquid flow, cfa

V, = saturated specific volume a t PI, CU fft/lb

1

The steam consumption

for a simplex double-acting pump and twice this for a duplex double-acting Pump. But

-

, '

Mechanical efficiencies and clearance ratios for reciprocating pumps

rate 0 ~ variation in reciprocating pump

tion stated earlier with equation (321, first estimate the steam cylinder terminal pressure, Pr, from

Emplo&

- INCHES

425

40

- WPLEX POWER

c-

TRIPLEX

AND

WMPS

MULTIPLEX

the factor C2 are given in curve form on Fig. 27. It will be noted that the condensation factor depends primarily (38) on the piston speed, but is also influenced by the steam cylinder diameter and the terminal pressure. The choice It is often useful to have this result in terms of piston of these three factors and the wgy in which they are used speed rather than capacity, and since is not accidental, but is based on an analysis of the heat exchange between the steam and the cylinder walls. WQ U = 2.6 Steam-End Valves and Linkages. The simplest (.~/4)D~'(l- S) (39) type of steam valve is the direct-acting D-slide valve showncan onbethe duplex pump of Fig. 23.where Thisthe type of for a simplex pump, substitution of this in the foregoing valve used only on a duplex pump valve steam consumption equation results in motion for one side is obtained from the motion of the W = 0.106 C1CZUDaZpt piston rod of the other side. For a simplex pump, a free(40) moving steam-operated main valve with a meahanically + or twice this for a duplex pump. actuated pilot valve is required since at slow speed there A further useful relation is obtained by dividing the might be only power enough to bring a mechanically Steam comum~tionby the liquid homepower. Thus the actuated main valve to dead center, thus causing the liquid horsepower (any fluid) pump to stop. Even for duplex pumps a slide valve canQP 144 not have lap or lead for the same reason. pw= A When a slide valve, either main or pilot, is used for a 550 vertical pymp, a lateral preseure is required to hold the valve against its seat; for this reason piston-type main valves are more suitable for vertical pumps. However, small-area slide pilot valves are found to be quite satis-

PW= 0.~03427Em(l - S)U(p, - p,)Da2

(41)

So the steam rate per fluid horsepower is

2(r/4)DWzL(1 S)

S.R.

=

w = 30.9 ClCZ -

p,

(42) Em(1 - 8 ) bt - pa) for either simplex or duplex pumps. Average values of the mechanical efficiency, slip ratio, and clearance ratio CI are shown in Fig. 26. Values of p w

Independent inlet and e&aust cyli&r ports are used with the direct-acting D-type slide valve-?n duplex pumps, the exhaust ports being on the inside so that as the piston approaches the end of its stroke the exhaust port is closed, and steam trapped ahead of the piston becomes compressed and serves as a cushion which prevents the piston from striking the cylinder head. For duplex steam cylinders of about 10 in. diameter and over with a direct-acting slide valve, a cushion-

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

MARINE ENGINEERING

release valve is usually fitted between the steam and exhau9t ports which is essentially an orifice the cushioning actionat the end of the stroke. The valve is located in a bypass between the inlet and exhaust ports and is manually adjusted by a handwheel on the outside. If the cushion valve is fully closed, complete cushioning is obtained; if the valve is partly opened, the compression is partly relieved by the amount steam flowing through the omhion valve. B~ manually adjusting the cushion valve a smooth stopat the end of the stroke can be obtained under all ordinary conditions of pressure and ~h~ application of a steamdperated main piston valve with a direct-acting pilot slide valve to a simplex pump is shown in ~ i 22.~ . this case the main slide valve of the fully balanced piston typeis controlled by the pilot valve and has a slide valve which rides above the

Slide valves are limited to sizes of about 25 in' pressures 150 psig' in diameter and steam of They are further to for diameters about 400 F for the larger sizes and 450 of 10 in. and kss. Above these temperatures it is difficult to lubricate the slide properly, and the likely to become warped and cause excessive leakage' Balanced piston valves are used when the size Or arc perature limits exceed those for which D-slide may bo suitable. For a duplex pump the piston must direct-operated, but for simp1ex pumps the consist, as previously explained, of a steam-operated main valve controlled by an auxiliaw of either the 'lido With a steam-operated main or piston type, starting and cushioning at the end of tho stroke are accomplished through the starting and ports. AS the main steam piston approaches the end its stroke, the main port, which is open to exhaust,

'*

In the case of a duplex pump it is evident that the slide valve must not reach dead center until the piston is near the end of its stroke and that this cannot hold for both pistons unless lost motion is provided in the valve linkage. Even with a liberal amount of lost motion some variation in the length of the stroke of a duplex pump is to be expected under different load conditions. 2.7 Liquid-End Valves. Two of the numerous types of liquid-end valves employed appear in the pump sectional illustrations. Stem-guided metal disk valves are ~houmin Figs. 22 and 23. Rubber composition disk valves may be used for moderate pressures handling cold water. For higher pressures \ring-guided valves with a conical facing are used. There are numerous other types of pump valves for various special applications. Among these may be menh n e d the spherical or ball valve and the semi-spherical valve, both desirable for handling viscous liquids because of the clear area through the seat, and the hinge or flap valve which is used for liquids carrying solid matter such HS sewage. It mill be noted that all valves are installed in 11 horizontal position and that the valves are all mounted

427

pumps are not suitable for operation with more than about 50 to 75 deg F of initial superheat, since a higher superheat prevents the formation of sufficient condensate to lubricate the moving parts. 2.10 Materials. For steam-ends, steel is used for greater strength in cylinder castings and heads when the steam pressure exceeds about 300 psig, and is used exclusively for this purpose in naval vessels because of the low shoclr resistance of cast iron. Ductile iron is also used and bas proven to be shock resistant. For pump ends handling fresh water, the cylinder castings may be of iron or steel depending on the pressure. The working parts generally are made of bronze; but piston rods, valve disks, and stems may be made of monel preferably or hewreated .atahless steel for greater strength or improved wearing quality.-'For m&umum corrosiorl resistance, one of several grades of austenitic stainless steel conforming to ASTM specific* tion A296 may be used. For seawater service, nonferrous working parts are used exclusively in the pump ends, and it is better t o have the complete pump end of nonferrous construction. Mineral oils of any character, including petroleum fuel

MARINE ENGINEERING

and lubricating oils, generally are handled by pumps of The pump-end cylinder liner and all-ferrous pistonhaterials are usually made the same as the comesteel is Bponding steam-eIld itemsbut hardened for valve seats, disks, and springs, with the valve seats of a material or hardness different from the Additional discussion of materials for reciprocating steam pumps and their design features are contained in references [9] and [161. 2.1 1 Applicatidns. At the present time Very few reciprocating steam pumps are used for continuous service. They are very useful, however, and continue to be used or emergency units for boiler feed service. as The simplicity of operation and self-priming characteristic of a reciprocating steampump make it well suited to bilge, ballat, general service, fuel oil transfer, and cargo oil stripping.

Vertical Pumps are preferred for thelarger sizes since less floor space is required for a given capacity- 4further advantage of the vertical Pump is that the weight the in less pistons is not carried by the cylinders, piston friction and wear. For the same capacity and Pressure) a simp1ex pump is generally lower in cost and more economical than the duplex type. A simplex pump is also more suitable for severe suction conditions where the pump may lose its supply of liquid temporarily. I n such m5es the simplex pump, which always operates on full stroke, may recover its suction sooner than a duplex pump which may shod stroke under these conditions. The steadier flow characteristics of a duplex pump, as compared to the simplex pump) have been mentioned' This effect may be offset to a considerable extent by the longer pause a t the end of the duplex Pump stroke-

Section 3 Power Pumps Basic Types. The term "power 3.1 Detlnitions pump" is used to describe any type of pump whose action depends on the recipmcating motion of pistons or plungers and whose motive power is from an external

The fixed-stroke power pump is driven through a crankshaft and connecting rods by a driving unit (USUally an electric motor) that is either direct connected, chained, belted, or geared to the crankshaft. Variablestroke power pumps usually employ other means of conVerting rotary motion of the applied power to the reciprocating motion of the pistons or plungers. Power pumps are further classified according to the arrangement of pistons or plungers. A single or simplex pump has one piston or plunger. A duplex pump has two pistons or plungers. In like manner a triplex pump has three, and a pump has more than three pistons or plungers. The reciprocating pistons or plungers may be single-acting or double-acting. 3.2

Fixed-Stroke Power Pumps

General Considerations and Applications. The pump end of a fixed-stroke power pump is identical in function uw side, .6.5 Rotary Compressors. A two-lobe-type of positive-displacement rotary air compressor resembles the pump shown in Fig. 33. The speed of this type of compressor makes it well suited for attachment to highspeed diesel engines to supply scavenging air. The principle of operation is identical with the lobe-type rotary pump described in Section 4. Other rotary pump types which are used successfully

444

MARINE ENGINEERING

as air compressors include the sliding-vane type and the liquid-sealed elliptical-casing fixed-vane type, similar to the vacuum priming pump described in Section 4. Rotary types generally are suitable for pressures up to 100 psig, but are predominantly used for lower pressures. Intermediate pressures up to 100 psig are usually furnished by reducing valves fitted in the 100psig (or higher pressure) system being served by a multi-stage reciprocating compressor. 6.6 Centrifugal Compressors. A motor-driven multistage centrifugal compressor of the tjrpe used for general service air supply is shown in Fig. 46. A noise-attenuating enclosure is usually provided, although such an enclo~ure is not shown in Fig. 46. Details of construetion, including the shape of impeller vanes, are quite similar to those of centrifugal pumps. The inherently high speeds of centrifugal compressors require them to be driven through a speed-increasing gear connected to the driving motor or engine, or they

may be direct connected to a turbine. Speeds in the range of 5000 to 40,000 rpm are common. They are used for diesel engine scavenging services and refrigeration and air conditioning services, as well as general ship air service. A high-speed turbine drive is ideal for special services requiring variable speed. Centrifugal compressors may be either single-stage or multi-stage, and they range up to 125 psig and 2500 cfm. A special advantage is their ability to deliver oil-free, clean air. They are essentially free of vibration, and noise can be attenuated to acceptable levels. The theory for centrifugal compressors is similar to, and is as fully developed as, the theory for centrifugal pumps described in Section 1. However, the compressibility of the gas is an additional factor which must be considered. A complete discussion of centrifugal compressor theory is given in references [3] and [6]. Special compressor applications are described in references (171, [18], and [25].

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS LIVE STEAM INLET

1

EXPANDING

*

SUCTION, INLET

CONVERGING

ABSOLUTE PRESSURE AT SUCTION

I

t

Section 7 Ejectors The steam jet ejector is employed principally to remove air and noncondens able gases fromvacuum equipment. The ejector is a simple type of compressor in which high-pressure steam is passed through a nozzle where it is expanded to a pres sure corresponding to the desired vacuum. The steam exits from the nozzle at a high velocity, and air and other noncondensable gases surrounding the issuing jet are entrained in the high-velocity jet and carried into a diffuser. The passage of the mixture through thii divergent tube effects conversion of the kinetic energy into pressure energy, thereby elevating the pressure a predetermined amount above the suction inlet pressure. The ejector, in schematic form, is illustrated in Fig. 47. Ejectors are classified as single or multi-stage depending on whether the compression is accomplished in one unit or several units in series. They are classified as single or multiple element depending on whether one or more units are installed in parallel. They are further classified as to being condensing or noncondensing, that is, with regard to whether or not the discharged steam is condensed. 7.2 Performance. The amount of pressure elevation, commo,nly referred to as compression ratio, that can be accomplished satisfactorily in a single stage of compree sion is governed by the initial steam pressure available for operating the ejector, the vacuum to be maintained, the importance of efficiency measured in terms of steam consumption, and the requirements for stability. Ejector stability under all normal operating conditions is a matter of prime importance. Unstable operation results when the steam jet momentarily and at irregular inter7.1

Classification and Types.

vals breaks down thereby allowing a temporary reversal of flow to take place through the combining tube; this can result in a partial or temporary loss of vacuum. The experience of the ejector designer dictates the maximum allowable ratio of compression per ejector stage that will ensure stability, after giving due consideration to the governing factors. For conditions usually encountered with marine installations, it is customary to employ single-stage ejectors for vacuum requirements up to 26+ in. of mercury, thereby establishing a maximum compression ratio of 8.57 to 1, and two-stage ejectors where the requirements exceed thii figure. Threestage ejectors can be designed having a better operating efficiency than two-stage ejectors for installations operating a t 29 in. vacuum or more, but they are seldom used in marine service by reason of their more. complicated structure, higher initial cost, greater weight and space requirements, and certain operating deficiencies, all of which combine to offset the initial advantage of better economy. An ejector must be designed for a predetermined minimum motive steam pressure. If the steam pressure a t the inlet to the ejector nozzles is less than the design pressure, the ejector will not operate satisfactorily and will not maintain a vacuum. An ejector will work satis factorily with a reasonably higher steam pressure than that for which it was designed, with a resulting increase in ejector steam consumption proportionate to the increase in absolute steam pressure above the design value. It is normal to operate ejectors up to 10 psig above rating to allow for steam line pressure fluctuations. Partial clogging of the nozzle has the same effect as reduced

445

Fig. 48

OISCHARBE Fig. 47

- INCHES

MERCURY

,

Capacity of two-stage ejector with inter- and after-condenser for various air sudion pressures and temperatures

Steam jet air ejector

steam pressure. Selection of the operating pressure is of great importance. Very low pressures are uneconomical, whereas extremely high pressures necessitate very small nozzle bores with consequent fouling difficulties. Operating pressures from 80 to 300 psig constitute a preferable range, with the upper limit being 150 psig for smallcapacity ejectors. Superheat is of no economic benefit; however, a nominal degree of superheat is desirable since it provides added insurance for a supply of dry steam to the ejector nozzles. Wet steam seriously interferes with the operation to an ejector, with the smaller sizes being affected more than the larger sizes. Wet steam also causes rapid erosion of the ejector nozzles and should be avoided by the use of suitable steam-line separators or other drainage devices located close to the ejector steam inlet connection. The air leakage into a system under vacuum is subject to considerable variation and does not lend itself to a precise assessment. Air leakage design values selected must be influenced by experience. The air removed from a steam condenser is saturated with water vapor. It is, therefore, necessary to determine the amount of water vapor contained in the air-vapor mixture to be removed from the condenser by the ejector. The water vapor component is dependent on the absolute pressure and temperature existing at the exit from the air-cooling section of the condenser or external air cooler. To provide ample air ejector capacity, the temperature of the air-vapor mixture at the air-vapor outlet of a welldesigned surface condenser equipped with either an internal or external air cooler is considered to be 73 deg F less than the saturated steam temperature corresponding

to the absolute pressure required at the steam inlet to the condenser. The amount of water vapor with which the air will be saturated can be calculated by the following expression: #

i-

where

W = pounds of water vapor per pound of air leakage pw = steam saturation pressure, inches of mercury absolute, corresponding to the assumed temperature of the air-vapor mixture pa = absolute pressure, inches of mercury, at the airvapor outlet of the condenser 0.62 = 18/29 = ratio of molecular weights of water vapor and air It is customary to provide sufficient ejector capacity to permit a condenser vacuum of 29 in. to be developed when thq cooling water temperature and loading conditions permit, even though a lower degree of vacuum is specified for the base design condition. For this reason the estimated ejector capacity is usually specified at an absolute pressure of 1 in. of mercury when intended for turbines. Figure 48 i&&-rates fypical performance curves for a two-stage ejector having intesYand aftercondensers for various suction inlet temperatures and pressures when handling a saturated air-vapor mixture. Figure 49 illustrates a similar set of c w e s applicable to a single-stage ejector. The steam consumption of air ejectors for a specified condition is dependent on design refinements and other highly important operating features; namely, the ability

446

0

Fig. 49

MARINE ENGINEERING

I 2 3 ABSOLUTE PRESSURE AT SUCTION

4

PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

5

- INCHES MERCURY

Capacity of single-stage ejector forvarious air suction pressures and temperatures

to handle an overload without an excessive loss of vacuum. r a ~ i devacuating ca~acitvfor starting UD and quick pick^* in case the"sysiem & accidentally flioded with air, and the ability to operate satisfactorily under conditions with verv low water auantities being: circulated through the inter- and after-condenser with due regard for pressure drop limitations at full power. For estimating approximately the steam consumption of air ejectors employed for marine service, a value of 4.5 lb of steam required per pound of air-vapor mixture removed at 1 in. of mercury absolute and 71.5 F temperature can be used for turbine service. For other lowvacuum services a value of 6.5 lb of steam required per poqnd of air-vapor mixture at 3.5 in. of mercury absolute and 113 F temperature may be used. 7.3 Applications. Air ejectors are used aboard ship for the following systems: Main propulsion engine condensing plant. Turbogenerator condensing plant. Low-pressure-type distilling plant condensers. Heating system drain condensers. Auxiliary condensers for condensing auxiliary steam. Air ejectors usually are installed in duplicate, one set being sufficient for normal operating requirements and the other set used for standby service and abnormal requirements. This arrangement is commonly termed a "twin unit" and is designed so that either one of the ejector elements may be inspected and cleaned without interfering with the operation of the other element. This may be accomplished by the use of interstage isolating valves or by compartmentation of the inter- and aftercondensers.

Fig. 50

TWO-stagedual-element air ejector

Figure 50 illustrates a twin two-stage air ejector unit with surface-type inter- and after-condensers arranged in a common shell employing interstage isolating valves. This type of unit is used extensively for condensers serving main propulsion turbines and for condensers of turbogenerator sets. The flow of steam and air-vapor mixture is depicted in Fig. 51. Live steam enters at (1) through strainer,,, (2) to second-stage nozzle (3) and also firstrstage nozzle (4). After expansion through the nozzles the mixture of entrained noncondensables and steam from the firststage nozzle enters the first-stage diffuser (9) wherein the first stage of compression to inter-condenser vacuum occurs. The compressed mixture then enters the intercondenser through interstage valve (5), wherein condensation of steam takes place as the mixture passes upward in successive vapor passes formed by longitudinal baffles (6). The remaining noncondensables exit through manifold valve (7) to the suction of the second-stage element (8) and are entrained by the jet of steam issuing from second-stage nozzle (3) and finally compressed to atmospheric pressure after passing through the secondstage diffuser tube (10). The discharge from the secondstage diffuser enters the lower section of the after-condenser through opening ( l l ) , and again passes upward in four successive vapor passes formed by horizontal baffles (12). Saturated noncondensables are expelled through the after-condenser vent opening (13) to the atmosphere.

.

I i 3

448

MARINE EN( PUMPS, FORCED-DRAFT BLOWERS, COMPRESSORS, AND EJECTORS

Condensate discharged by the condenser condensate pump is circulated first through the inter-condenser and then through the after-condenser in a closed circuit; the condensate flows inside the tubes, thereby preventing absorption of noncondensables by the condensing medium while effecting full heat recovery through condensation of the ejector exhaust steam. The drainage from the inter-condenser, which operates at a vacuum somewhat lower than is maintained in the condenser being served by the ejector, is effected through opening (14), by means of a loop seal drain pipe connected to the condenser. The amount of loop seal required depends on the maximum vacuum differential. A minimum loop seal of 7 ft is usually sufficient for most ejector installations. The drainage from the after-condenser dhrough opening (16) is usually led by gravity to a freshwater collecting tank located in the engine room. Although air ejectors can be designed t,o operate without condensers, all marine ejectors are condensing in order to conserve feedwater and recover heat. The use of an inter-condenser between stages of a two-stage ejector effects a substantial saving in overall steam consumption, since the motive steam of the first stage plus some of the entrained vaDor are condensed in the inter-condenser. thereby greatly reducing the required second-stage capacity. A turbine gland steam condensing section is frequently provided in the air ejector after-condenser, thereby effecting additional heat recovery (by the condensate being used as circulating water) through condensation of the turbine gland leakoff steam. When such an arrangement is employed, a gland leakoff exhauster is connected to the after-condenser vent for maintaining a slight suction effect in the after-condenser. The slight vacuum must be sufficient to induce flow of the gland leakoff steam through suitable pipes leading from the turbine glands to the air ejector condenser. A suction effect of 5 to 10 in. of water usually is sufficient for this purpose. In order to cool effectively the noncondensables prior to their exit from the after-condenser vent opening, it is preferable to limit t,he condensate circulating water outlet temperature from the air ejector condenser to a maximum of 140 F. When insufficient condensate is available to maintain this limiting temperature, due to light loading or while maneuvering, recirculation is employed which may be manually or thermostatically controlled. Air ejectors for main engines and turbogenerators always use condensate as the cooling medium. Ejectors for distiIling plants use the saltwater feed as a cooling medium, thereby providing a stage of feedwater heating. The ejector principle is used also to pump water and other fluids against a low pressure by means of a highpressure jet. They are used aboard ship for laundry and plumbing system drains and are commonly referred to as "eductors." They offer a convenient means of evacuating remote compartments that are not connected to a drainage system, but which have a seawater fire main nearby that can furnish the motive water. They may- be motivated by air pressure and often are combined with centrifuga1 pumps to afford a means of priming.

Further discussion of design and operating features of ejectors is contained in references [4], [5], and [26]. References

1 S. A. Moss, C. A. Smith, and W. R. Foote, "Energy Transfer Between a Fluid and a Rotor for Pump and Turbine Machinery," Trans. ASME, 1942. 2 A. Stodola, Steam and Gas Turbines, translated by L. C. Loewenstein, Peter Smith, New York, 1945. 3 George F. Wislicenus, Fluid dfechanics of Turbomachinery, Dover Publications, Inc., New York, 1965. 4 I. J. Karassik and R. Carter, Centrzfugal Pumps, F. W. Dodge Corp., New York, 1960. 5 I. J. Karassik, Engineers' Guide to Centrifugal Pumps, .McGraw-Hill Book Company, Inc., New York, 1964. 6 G. T. Csanady, Theory of Turbomchines, McGraw-Hill Book Company, Inc., New York, 1964. 7 S. Lazarkiewicz and A. T. Troskolanski, Impeller Pumps, Pergamon Press, New York, 1965. 8 F. J. Wiesner, "A Review of Slip Factors for Centrifugal Impellers," Trans. ASME, 1967. 9 Hydraulic Institute Standards, Hydraulic Institute, New York. 10 A. W. Feck and J. 0 . Sommerholder, "Cargo Pumping in Modern Tankers and Bulk Carriers," Marine Technology, July 1967. 11 G. F. Wislicenus, R. M. Watson, and I. J. I Up, as required by equation (41). The methods and procedures outlined represent a reasonably accurate method for evaluathg condenser design geometry and performance. Accuracy can be improved by a more rigorous treatment of the fluid flow chaxacteristics. The degree of accuracy improvement necessary depends on the circumstances of a particular case. While the descriptions presented have been related to the normal geometry of the larger main condensers having two separate and distinc$kube bundles, the method is also applicable to single-bundle condensers such as those generally used for auxiliary condensers. The principles can also be adapted to condensens having more than two separate tube bundles or to any shell-and-tube apparatus which condenses a vapor on the shell side. References

where

1 Courtesy, Foster Wheeler Corporation. 2 Courtesy, Worthington Corporation. R r = -1= overall resistance frdm rational equa3 Courtesy, Heat Exchange Institute. tions, hr-sq ft-deg F/Btu ur 4 Courtesy, American Society of Mechanical Engi1 overall resistance from fluid flow char- neer, R ~m = -u,, = acteristics, hr-sq ft-deg F/Btu 5 J. E. Fowler and R. E Brandon, "Steam Flow DisIf fouling resistance is introduced (it must be con- tribution at the Exhaust of Large 'steam Turbines," lridered to be the same for both the resistance calculated ASME paper 54SA-62. by rational means and by fluid flow characteristics) equa6 J. P.&bald and W. D. Nobles, "Control of Tube tion (42) becomes Vibration in Steam Condensers," Proceedings of the American Power Cmference, ~llinbis1nstitute7of ~ e c h (43) nology, Chicago, Illinois, 1962. t~ndmay also be written as 7 -A. P. ~ i l b u r n "Problems , in Design and Research on Condensers of ~ a ~ o u i k T a nVapo3 d MixtgesJJJThe ' Rdm = (R, ry) - (Rr rf) (44) Institution of Mechanical Engineers, 1951. 8 "Standards for Steam Surface Condensers," Heat where Rdm= design margin as a resistance, hr-sq ft-deg lp/Btu, and must be positive if the condenser is to meet Exchange Institute, New York. 9 W. H. McAdams, Heat Transmission, McGrawi t ~required l performance. The performance factor PF may be expressed either in Hill Book Co., New York, 1954.

+

+

t

HEAT EXCHANGERS

489

CHAPTER X I V

Charles D. Rose Philip Liu

I Heat Exchangers Section 1

1.1 General. Heat exchangers are used throughout the marine power plant to transfer heat from one fluid (liquid or gas) to another fluid. The most widely used type of exchanger in marine service is the "shell-andtube" type. As shown by Fig. 1, the shell-and-tube heat exchanger consists of six basic elements: the bonnet, tube sheet, shell, tubes, baffles or support plates, and tie rods. The bonnet or channel is often referred to in the marine industry as the "head" or "waterbox" of the exchanger. Due to stringent space limitations imposed on marine heat exchangers and for ease of maintenance and cleaning, head inlet and outlet connections are arranged to permit access to the tubes and tube sheets without dismantling the attached piping. Only the smaller heat exchangers (under 100 lb) would have straightaway connections as depicted in Fig. 1. Figure 1 illustrates a "single-pass" fixed tube sheet exchanger. "Single-pass" is a term indicating that the tube-side fluid flows in one direction only. A "two-pass" or "four-pass" exchanger would have the tube-side fluid inlet and outlet connections a t the same end. The fixed tube sheet construction depicted in Fig. 1 (tube sheets welded to the shell) must often incorporate an expansion joint in the shell to compensate for differential expansion between the shell and tubes as a result of the relative temperatures of the fluids involved. The various means of providing for differential expansion are covered in Section 1.2. I n addition to fixed tube sheet designs, heat exchangers employing floating tube sheet and "U"tube designs are commonly used. Figure 2 illustrates the floating tube sheet construction which is employed in main lube oil coolers and electronic equipment coolers. In this design, one tube sheet is free to "float" against packing rings (usually made of neoprene). The packing rings are held in place between the head and shell flanges by a packing retainer ring. Details of this construction are described in Section 3.1 of this chapter. Zinc protectors, such as those shown adjacent to the division plate in the head of the heat exchanger in Fig. 2, are used in saltwater-cooled heat exchangers to protect headers, tube sheets, and tubes from galvanic corrosion by electrolytic action. When dissimilar metals that are connected together are immersed in an electrolyte, a simple galvanic cell is formed and an electric current

BONNET

TUBE SHEET

BAFFLE OR SUPPORT PLATE

TIEROD

TUB~SIDE NOZZLE Fig. 1

Typical single-pass conventional exchanger (flxed tube sheet)

flows from one metal to the other through the electrolyte which completes the circuit between the two dissimilar metals. The metal (anode) from which the current flows will tend to suffer rapid corrosion, often termed galvanic or electrolytic corrosion, and the metal (cathode) to which the current flows will tend to be protected from galvanic corrosion. The direction in which the current flows depends on the composition of the metals or alloys exposed to the electrolyte and also on the hardness of the metal, the cleanliness of the metal surfaces, and other factors. Thus, if a single metal is immersed in an electrolyte and one part of the metal surface is harder than another part, or cleaner than another part, there will be a flow of current from one part to the other and galvanic corrosion will take place. If several different metals or alloys are involved, current will flow in varying proportions between the surfaces exposed to the electrolyte. If clean metallic zinc is properly arranged within a heat exchanger waterbox, a current will tend to flow from the zinc to the adjacent metal surfaces exposed to the seawater which constitutes the electrolyte of the galvanic cell. The zinc protector plates are corroded as the current is generated; and the current flowing through the seawater to the metal surfaces of the heat exchanger tubes, tube sheets, and waterbox tends to protect these parts from galvanic corrosion. The elechc circuit is completed through the metal parts of the heat exchanger. Gaskets between the waterbox and manhole covers, to which zinc protectors are frequently attached, and

Fig. 2

Typical heat exchanger with floating tube sheet, disassembled

between the tube sheets and the water chests do not interrupt the flow of current, as the circuit is completed through the metallic bolts and collar studs which secure the joints. The Navy requires that the exposed surface of sincs (exclusive of edges) be a t lea& one square foot for each 1000 square feet of heat transfer surface. For now construction, the amount of zinc surface is based on rtn equation which is given in Military Specification MIL-A-19521. All zinc protectors should be thoroughly scaled once every 4 to 6 weeks to assure that active metallic zinc nurface, as opposed to corrosion scale adhering to the metal, is exposed to the seawater. Zinc surfaces which aflord proper protection are quickly detefiorated. In the design of the waterbox and application of zinc modes, care must be exercised that the zincs do not interfere with, or add turbulence to, the fluid flow within the waterbox. As illustrated in Fig. 1, the shell of the heat exchanger is usually cylindrical with flanges attached to each end. The tube sheets are either welded or bolted to the shell flange, and the heads are bolted to the tube sheets. The most critical joint in the exchanger, and that most likely to develop a leak, is the tube to tube sheet joint.

I

SERRATIONS Fig.

9

Expanded tube-to-tube-sheet joint with inlet end bell

Expanding the tube into the tube sheet with a "tube expander" is the most commonly em~loyedpractice; a typical expanded tube is &wn by Fig. 3. Care must be taken to follow the manufacturer's instructions as to limits of tube wall reduction and dial settings on the electrically or electronically controlled automatic tube expander when repairing a leaking tube to tube sheet joint. Seawater flowing into heat exchanger tubes a t high velocities tends t9 remove the thin protective film of

MARINE ENGINEERING

HEAT EXCHANGERS Table 1

T* Fluid

ITEM side

Velocity, fps

LWBICA~G OIL COOLER

Low-paEssmr~1 F ~ HEATER D

HIGH-PEEBSUBE FEEDHEATEB

Seawater

Feed

0-7

4-7 2-8

Typical Featuror

-

TANKCLEANING SYBTEM --HEATER DRAINCOOLER

5/811 OD

Tube Material H a d Material

per inch 90/10 CuNi Bronze

90/10 CuNi Steel

70/30 CuNi Steel

90/10 CuNi Steel

Tube Sheet

90/10 CuNi

Steel

Forged Steel

Bronze--Composition G

SluU Side

Fluid Velocity, f s Number 6 P Pas= Type of b d e

Flow

Cbnslrvdion

5/8" OD

x 0.049"

Steam Condensing 6 (baaed)

Oil

2-3

10-16 (baaed)

Expanded

m d e d

Expanded or welded

Thermal Ekpansion Gasket

Floating tube sheet Full Face

Expausion joint

U-Tubes

corrosion products adhering to the base metal of the tube wall. This protective a m is replaced a t the expense of further corrosion of the tube wall. As continued removal and replacement of the protective film of corrosion proceeds, the tube wall is gradually thinned and the tube to tube sheet joint is weakened and ultimately fails or the tube wall just beyond the tube sheet is perforated. This type of erosion is generally termed impingement erosion, inlet end attack, air erosion, or bubble attack. The occurrence and rapidity of the attack are governed by the water velocity (for recoinmendations see Table 1)) the amount of air entrained, and the design of the waterbox as it affect8 the velocity, ,direction, and turbulence of the fluid flow approaching the tubes. The inlet ends of the tubes are normally ground flush with the face of the tube sheets, and no gaps should be left between the edges of tubes and the radius of the holes on the inlet sections as gaps here tend to promote impingement erosion. The outlet ends of tubes may

Collar Bolts Bare

600

2.7 gpm seawater per sq ft

190°F 110°F rise in

200 180 lb drains per hr/sq ft

100°F 200°F drop in d r u i ~ ~ ~

extend Ne in. beyond the face of the tube sheet. Where s considerable temperature differential exists withim a heat exchanger, a packed-tube design may be employed. Packed tubes, such as shown by Fig. 4 and Fig. 8(c), allow considerable differential tube expansion since each tube is free to move independent of the others. The combined low-pressure feed heater/drain cooler/gland-exhaust condenser (which is described in Section 3.5) is a typical marine exchanger normally employing packed tube ends. Marine heat exchangers are designed in accordance with the Standards of the Tubulas Exchanger Manufacturers Association [I]' and the American Society of Mechanical Engineers [2] in addition to compliance with marine regulatory body wde requirements such as Lloyd's [3], U.S. Coast Guard [4], and the American Bureau of Shipping [5]. Many additional standard 1

Condensing

2

CuNi

I

{ lrrt~dsteam !rb~dst~sing fi rlr H (baffled)

Full Face

Ring (shell aide) Solid co per ring (tube d e ) Stud Bolts Bsre

10°F 1 0 0 9 rise in feed

i'i (bafled)

Steam

U-Tubes

Ring or flexible

800 3 0 0 lb feed per hr per sq f t

Heater dr-ains Segmental Counterflow, last pwa

...

Tube Joint

Bolting Collar bolts Surfaca Finned Therm41 PerfOnna?ca Heat T r a d e r Coefficients ~tu/hr-fd 40 0.3 to 0.4 gpm L.O. Surfaca Requirement per sq ft Terminal Temperature Diffknce (or LMTD) 18°F Temoerature 20°F d r o ~in L.0.

Steam Condensing 6 (baaed)

Segmental

Segmental counterflow

ELECTBONICB EQUIPMENT COOLER

Saltwater 0-7 1-4 3/4" OD x 0.049"

1 R ~ 2 ~ 5/8" 7lowODfin-19x 0.049" fins

x 0.049"

w1 Marine Heat Exchangers

Numbers in brackets designate References at end of chapter.

8;;h l t a

Steam

steam'

Condensing

Condensing

2

2

(bayonet tube) 90/10 CuNi Steel Steel or Cast Steel 90/10 CuNi BronzeSteel Composition G or steel Fresh Water

Chilled Water or Seawater

Boiler Makeup Water

1

:;"

34

5/8" OD X 0.049 (bayonet tube) BWG ,. Steel 90/10 CuNY Cast Steel Bronze 4

Water

1/4" OD X 0.035"

4-7 2 5/8" OD XO.049" BWG 90/10 CuNl 90/10 CUNI

/

Steel

90/10 CuNi or Bronm Composition G

...

90/10 CuNi

Steam Condensing

...

Open Shell

Lubricating Oil 2-3 1 (baf8ed)

Fuel Oil 1-3 1 (baffled)

Fresh Water 3-5 1 (baffled)

Seawater 1-8 None

None Boiling

Segmental Counterflow

Segmental Counterflow '

Segmental Counterflow

None , Counterflow

Expanded

Expanded

Floating head or U-Tube , Full Face

Bayonet

Expanded Expanded (buter) Ferrules (inner) Bayonet ...

Stud Bolts Bare

Stud Bolts Finned

10-20 60 6-8 gph 75 gph

1 (baffled)

Welded or brazed

...

U-Tubes

Ring

Full Face

Ring

Stud Bolts Finned

Stud Bolts Bare

Stud Bolts Bare

Stud Bolts Bare

45 7-12 gph per sq ft

400 150 h per Sq t

500 0.25 m boXr water per sq ft

150 hper sq%

100°F 100-150°F

20°F 5°F drop in fresh water

400°F 400°F drop in sample temp.

200°F 80°F rise in water temp.

Ring

, ,= 6-10 111 team

&r Irr/sq ft

8% nonucr~~tlansable

%-50 lb steam per hr/sq ft

80°F Boiling a t 100 pa aba

P Y = #?sq 120°F

250°F 50°F

T

500

Heads are designed so that it will not be necessary to disassemble piping to gain access to the inside of the heads The and tubes. tube bundle is usually of the removable type for easy cleaning and maintenance. I n the design of cooling water spaces and connections, a smooth flow path must be provided to minimize erosion-corrosion attack. Sharp comers and projecting edges are avoided. Internal fittings are arranged to result in a minimum of interference with the water flow and a minimum of turbulend? Y ' Rg. 4 Packed tube Cooling water velocities a t the design point must not exceed those specified or recommended by the material supplier. requirements have been incorporated in specifications for Heat exchangers having tubes of length exceeding marine heat exchangers due to stringent space limitations 4 f t are designed so that the o r d e ~ n glength of tubes and reliability requirements. The following points are will be in multiples of 6 in. The ordering length of tubes is determined by addbg M in. to the face-to-face emphasized in the design of marine heat exchangers:

MARINE ENGINEERING

HEAT EXCHANGERS

(g) Head integral with shell, no shell side

(a) Flat face ring shell flange with full face unconfined gaskets on both shell and head side.

(d) Welded shell tube sheet joint with semiconfined gasket on head side.

gasket.

support plates or between a tube ahget and a support plate does not exceed 36 in. Holes for tubes in bases and support plates, baffle clearances, and tie rod standards are usually required to be in accordance with the latest standards of the Tubular llxchanger Manufacturers Association [I]. In order to diffuse the entering stream and reduce erosion of tube ends, for single-pass coolers the waterbox depth measured normal to ihe-tube sheet should be not less than one-half the equivalent diameter of the tube sheet area exposed to the flow of the cooling water into the tubes. For cylindrical two-pass coolers, the waterbox head depth ~houldbe not less than 35 percent of the inside shell diameter. All heat exchangers must be provided with tuiequate foundation supports. When required by the oonditions of service, provision is made in the design of the supports to provide for expansion or contraction of the shell. Heat exchanger supports are usually independent of any att,ached piping. Supports must be given special consideration when designing for high-impact shock conditions [6]. Is the design of marine heat exchangers, consideration must be mven to the varvin~deerees of inclination encountered in service. In n a v z prLtice, heaters and ooolers are designed to perform satisfactorily under oonditions of 5 degrees trim, 10 degrees pitch, 15 degrees list, and 45 degrees roll (in commercial practice a 30degree roll is the design criterion). The conditions for permanent list and roll or for trim and pitch are not bonsidered additive. Adequate air vents must be provided on heat exchanger waterboxes to avoid the collection of air in the upper region of the waterbox, as air pockets can restrict the tube-side flow and render a portion of the heat transfer surface ineffective. Such air pockets can also result in overheating and expansion of the dry tubes and oause failures of the tube joints at the tube sheets. In feedwater heaters and condfnsers, wet steam at a high velocity must not be permitted to impinge on the tubes, otherwise the surface of the tubes will be rapidly eroded. ~ a f h e sor distribution pipes must be incorporated as necessary to preveht the direct iinpingement of wet steam on the tubes.

n , 00000 000000 0000000 000000 Seg,,n*l

(e) Double packed floating tube sheet with retaining ring and packing rings on both hell and head side.

(h) Double floating packed tube sheet with retaining ring and packing rings on both the shell and head sides.

7

(c) Ring shell flange with tongue and groove joint and fully confined flat ring gasket on both the shell and head side.

If] Outside-packed removable tube bundle with semiconfined gasket on head side. Fig. 5

(I) Ring joint flange, double tube sheet design with confined O-ring gaskets.

Shell, tube sheet, and heod ioint design

distance between the outside faces of the tube sheets. The minimum tube sheet thickness is usually specified to be not less than 31 in. When external fins (low fins) are applied to tubes, one end of the tube is usually gradually enlarged to the outside diameter of the fins to enable the removal and insertion of individual tubes. Holes in the tube sheet at the inlet end of the tubes are flared to allow for belling the ends of the tubes. Holes in the tube sheets are provided with at least one groove. The edges of the holes are rounded, usually on a

He-in. radius, on the inner face of each tube sheet and on the outer faces of the tube sheets at the discharge ends of the tubes. The inlet ends of the tubes are expanded and belled and the ends are finished flush with the face of the tube sheet. In no case should the ends of tubes be below the face of the tube sheets. Discharge ends of tubes can protrude up to He in. beyond the face of the tube sheet. a A number of the baffles are increased in thickness (usually H in.) to act as tube support plates and are located so that the maximum tube length between

1.2 Shell, Tube Shed, and Head (Channel) Joints.

The design of the means for attaching the head (channel), tube sheet, and shell are governed primarily by the operating pressures and operating 'temperatures to be ttccommodated. Figure 5(a), ' which employs gaskets, depicts the least costly means. Gaskets (which are usually compressed asbestos) aiqepositively positioned, thereby insuring alignment. This design could incorporate studded tube sheets (shown) or collar studs to permit removal of the head without disturbing the shellside pressure joint. In naval applications, the joint shown in Fig. 5(a) may be used with up to 150 psig and 375 F. Figure 5(b) illustrates the joint that is the most widely used in shell, tube sheet, and head joint attachments.

BaMe

(dl Triple Se~nental

P

(bl Single W m n t e l

-

-

(b) Flat face ring hell flange with ring type unconfined geskets on both shell and head side.

, l"

.

,

'(el D i x 81 Doughnut

(c) Double Segmental fig. 6

Fbw baffler

The design normally incorporates studded tube sheets (shown) or collar studs. The recommended gasket material L coinpressed asbestos, and the usual a plications will accommodate liquids and vapors to 3 psig and 450 F. The design indicated in Fig. 5(c) is widely used for high-pressure fuel oil and steam service. The gasket grooves completely confine and positively align pass partitions, and afford excellent protection against gasket blowouts and failures. The usual applications permit ' liquids and vapors to 1000 psig and 750 F with a proper gasket material and design (normally compressed .asbestos, jacketed asbestos, or spiral-wound gaskets). Figure 5(d) depicts a joint that is used primarily when no leakage is permitted o p h e shell &de, as in cases where the shell-side medium is either hazardous er Comsive. The semi-confined gasket allows for protection from blowing out and provides a more positive positioning of the gasket. The usual applications permit liquids and vapors to 600 psig and 500 F with proper gasket materials (normally compressed asbestos, jacketed asbestos, or spiral-wound gaskets). f i e design indicated in Fig. 5(e) is an excellent means

L

494

HEAT EXCHANGMS

MARINE ENGINEERING

(4Square tube pitch Fig. 7 ' Tube pitch

(a) Triangular tube pitch

Id) Floating head with backing device

{b) Bayonlrt Tube Fig 8

(d) Rotated square tube pitch

(b) Rotated Triangular tube pitch

ggmpatible fluids; should a tube joint fail, the leak can be detected immediately, avoiding contamination of the guids. The usual applications permit liquid and vapor prassures up to 1000 psig and 400 3 '. 1.3 Spseifle Construction Detuils. I n shell-andCube heat exchangers, baffles are generally used to guide flow and increase the velocity of the fluid flowing on the rrhell side of the heat exchanger. The most commonly uaed baffles are of the segmental type or its variations, as Illustrated by Fig. 6. Segmental baffles are formed by cutting out thin metal plates to an outside diameter slightly less than the b i d e diameter of the shell. A segment is cut out of the baffle to form a segmental opening, the size of which may vary from approximately 15 to 45 percent of the nhell cross-sectional area (with single segmental, double regmental, disk, and doughnut ba&) or higher than 46 percent of the shell area (with triple segmental baffles). The tubes must be supported at intervals dong their length to minimize tube vibrations excited by the fluid flowing across the tube or by pulsations of the flow rate. The maximum permissible unsupported tube longth will depend on the tube material and tube diameter. For nonpulsating flow, the maximum unsupported tube length is 60 to 80 times the tube diameter. For pulsating flow, the unsupported length is uonsiderably less, and welding or brazing is used in lieu of spacers and nuts to secure the tube support plates. When designing for high shock, the tube support spacing $

of providing for shell or tube thermal expansion. When the packing on one side becomes deteriorated and eventually develops a leak, the leak is readily detectable through the vent and drain holes in the retaining ring and the packing can be replaced. The studded retainer ring permits repacking the tube side with full p r e m maintained on the shell side. The usual applications permit liquids axid vapors to 300 psig and 300 F (the temperature is limited by the packing material). The design indicated in Fig. 5 0 is an alternative means of providing for shell or tube t h e d expansion. On the she11 side, liquids and vapors at pressures up to 500 paig can be accommodated but the temperature is limited (usually to about 300 F) by the type of packing. On the tube side, the pressure may reach 600 psig at temp&rattwesof 500 F with proper gaaket materials. Figure 5(g) describes a joint which is relatively inexpensive and is used when it ia desirable to eliminate the shell-rside gasket. This design employs only one

gaaketed joipt at each head end for servicing the tubes. There are no limits on pressure or temperature except for gasket considerations. The design indicated in Fig. 5th) is recommended where contamination of one fluid medium by the other cannot be tolerated and provides for excellent thermal ex~ansionof either the shell or tubes. The studded r e k n e r ring permits repacking the tube side with full pressure maintained on the shell side. The double tube sheet type of construction is often specified by the Navy in cases where the greatest assurance against mixing of the two fluids is desired. Liquids and vapors to 300 psig and 300 F (with due consideration to packing material) can be accommodated with this design. The joint described by Fig. 5(23 is excellent for vacuum and very-high-pressure service, and the temperature iR limited only by the O-ring gasket material. The design requires precise machining of the O-ring grooves. Tho double tube sheet design-is excellent for use with non-

T h m o l expomlon provisions

is made small to minimize thd%esponseof the tubes to shock loadings. There is considerable latitude in selecting the pitch for tubes. The four most common tube patterns, as viewed from the tube sheet end,, am shown in Kg. 7. The triangular pitch and rotated triangular pitch are the most compact forms, and the triangular pattern is the one most commonly used for marine heat exchangers. Square pitch and rotated-square pitch pattern have see-through lanes which facilitate manual or mechanical cleaning of the outside of the tubes. The square pitch is common in submerged-tube boilers where passage for fluid circulation is important. It is a h applied in s e ~ c e whme s minimum pressure drop is a paramount design criterion. Tube patterns other than' the four illuatrated in Fig. 7 could be used to dtisfy specific design co%iderations as to pressure drop or turbulence, but they would be more costly to manufacture. Tube center-to-center distances are normally 6.25 times the tube diameter, or greater, and uniform over the tube field. .*@ Since the shell and tubes operate at&fFerenbtemperatures, it is n q s a r y to provide meam to accommodate the diflerence in their thermal expansion, as h i stresses could otherwise be developed within the heat exchanger. The various types of construction which have been used to accommodate thermal expansion are illustrated in Fig.5(e), Cf), and (h) and Fig. 8; each has its own area of application.

I

496

MARINE ENGINEERING

HEAT EXCHANGERS

expansions present no difficulties with this type of design. When the tubes are firmly fastened to the tube sheets and the tube sheets are fastened to the shell (i.e., fixed tube sheet designs), shell expansion joints are used to lessen the stresses caused by the difference in thermal expansion between the shell and tubes. The shell expansion joints illustrated in Fig. 9 are adequate only for small differential thermal expansions (say 0.06 in. for a 5-ft tube). For large thermal expansions, expansion joints of the bellows type are normally used. To provide for thermal expansion when the shell of a heat exchanger operates at an elevated temperature, (a) Shell expansion joint using flanged only one "foot" or "leg" of the shell is anchored to the and flued heads foundation and the other is designed to permit free axial movement sg that the axial expansion of the shell is not resisted. Freedom for axial movement may be provided by either a sliding foot (slotted holes with ferrules) or a slender leg (in the case of heavy heat exchangers) which provides little resistance to lateral force. 1.4 Design Data Requirements. In order to specifically direct attention to the items which govern the design of a heat exchanger and which must be furnished to the design engineer (or assumed by him), those items which should be included in a specification for a marine heat exchanger are listed below. Substance to be heated (and cooled) (b) Split pipe shell expamion joint Quantity of substance to be heated (and cooled) within a given period of time Fig. 9 Shell expansion iainh Initial temperature of the substance heated (and cooled) Final temp rature desired for the substance heated The simple U-tube illustrated by Fig. 8(a) is the most (and cooled) economical and commonly used means to separate the When the heating and cooling media are other than thermal expansion in the tubes from that of the shell. water or commonly known substances, the following I t is widely used in small condensers and instantaneous should be specified : heaters. (a) Viscosity The bayonet tube, illustrated in Fig. 8(b), is suitable (b) Specific gravity for use with a tight, cross-flow baffle spacing. Bayonet ' (c) Specific heat tubes are used in viscous oil heaters. (d) Thermal conductivity Packed-tube joints, Figs. 4 and 8(c), permit some axial Working pressure of the heated (and cooled) movement of the tube. Packed-tube joints are used in substance rectangular (box type) low-pressure feed heaters and in Allowable pressure drop through the shell and tube smaller straight-tube condensers. sides of the heat exchanger In cases where a large differential thermal expansion Desired construction materials can be expected, a floating head arrangement may be Typical features of a variety of marine heat exchangers employed. As can be seen from Fig. 8(d), large thermal are listed in Table 1.

T9

t = temperature of cold fluid A, = tube outside surface area 2~ = rio rdio rw rdo ro r, = resistance across the fluid film on the inside of the tube rdh = resistance of the deposit or scale on the tube inside wall r, = resistance of tube wall metal rdo = resistance of the deposit or scale on the tube outside wall ro = resistance of the fluid film on the outside of the tube

+ + + +

1i I

i

!

.

Section 2 Heat Transfer in Shell-and-Tube Heat Exchangers Q =

Ao(T - t) Zr

where

Q

(1)

I1 I

= heat transferred

T = temperaturk of hot fluid

SCALEPR DIRT

Fig. 10 Heat transfer through a tube

b

2.1 Heat Transfer Relationships. The fundamental theory dealing with heat transfer was discussed in Chapter 2, and a discussion of heat transfer by condensation is given in Chapter 13. The application of this theory to the transfer of heat through the walls of a tube leads to the equation

TUBE ?ALL

These quantities are illustrated by Fig. 10, and the oorresponding temperature gradients across a tube are ahown in Fig. 11. I t may be noted that "new clean-tube oxide film resistances" have not been included on the inside and outside of the tubes as they are not of sufficient magnitude to warrant consideration in the design of viscous fluid or water-to-water heat exchangers. The film resistances in equation (1) are further defined

as 1 1 1 ri0 = =(2) hiAi/Ao hi, tube-side film coefficient 1 1 + To=-- (3) h. shell-side film coefficient A, and A; are the tube outside and inside surface areas respectively. The terms hi and ho are called specific conductances or heat transfer film coefficients and are generally a function of flow velocities and fluid properties. The scale resistance or fouling resistance terms, rdo u,nd rdo,do not lend themselves to analytical assessment and little is known regarding them. The resistance of the tube wall metal is readily analyzed and it can be quantified without difficulty. The overall heat transfer coefficient, U, which is defined as the reciprocal of Zr, is a convenient means of axpressing the resistance to heat transfer through tubes. The overall heat transfer coefficient can be stated as

Fig. 11

Temperature gradients auou'tubes

location within the exchanger. Consequently, the heat flow rate, Q, also varies from location to location. To obtain the total heat transferred per unit time for a heat exchanger, it is convenient to base calculations on the average temperature difference across the tube over the entire tube length. Such a temperature difference is defined as the log mean temperature difference (LMTD) and is expressed mathematically aa LMTD =

-1

1 .A0

(T - t)dA

(5)

By combining equation (5) with equation (1) and noting that Zr is equal to the reciprocal of U: the heat transfer through a tube is determined to be

Equation (6) is the basic analytical tool employed to establish the thermal design of a heat exchanger. The factors to be considered in assessing this equation are briefly discussed in the following. With known terminal temperature differences between the shell and tube streams, the mean temperature difference between the shell 'and tube flows can be derived [7] if the following aasumptions are made: The overall heat transfer coefficient is constant along the entire flow path. The flow rate and.pxific heaj of both the shell and tube streams are constant. The heat transfer surface is uniformly distributed along the flow paths. The temperature of either fluid is constant over 1 any cross section of its path (i.e., there is complete U = (4) (I/&) rdo rw rdio (Jlhio) mixing and no stratification). There is no internal leakage or bypassing of fluid In the case of shell-and-tube heat exchangers, the temperature difference (T - t) varies from location to around the tube bundle.

+ + + +

498-

MARINE ENGINEERING

499

HEAT EXCHANGERS

for C allows a considerable m a r k (as com~aredwith a value of 0.33) to allow for flowleakages asLs-would occur tube-side film coefficient between cross baffles and the shell, and should only be tube inside diameter used as an average value. thermal conductivity of fluid on tube side The term "ideal tube bank" came from a report [lo] 0.027 published as s result of a research program on shell-andtube-side mass velocity tube heat exchange? conducted at the University of tube-side fluid viscosity at bulk temperature Delaware. The research /program was supported by tube-side fluid viscosity at tube wall temperature various heat exchanger manufacturers and lasted more Prandtl number = CJ/K than 12 years,, An "ideal tube bank" came to mean a specific heat rectangular hube field with straight-through flow and no Figure 13 is a nomograph of the film coefficient inside wall effects or bypassing. Theresults from these tests tube walls for turbulent water flow inside No. 18 BWG can be presented in the form of equation (13) if the contubes as determined from equation (12). A means of stants C and m are replaced by functions of the tube oorrecting for other tube dimensions is indicated in pitch, P, the tube diameter, do, and the Reynolds number, (doG/Z), in the lower Reynolds number range. Fig. 13. The relationship used to assess the shell-side film That is C = b(1.33d0/P)" (14) ooefficient is similar in form to that used on the tube where aide and is as follows:

where

(c) T w singbshell and double-tube pass exchangers in series

Q) Ona shall pass and multiple of twa tube passer

RQ. 12 Flow arrongeaenh Light flow Pnes mpreaent lube-side flow and heavy itow lines represent shd-ride flow

There is no transfer of heat between the heat exchanger and its surroundings. For a single-pass counterflow heat exchanger such as illustrated by Fig. 12(a), the LMTD is

Single-shell heat exchangers with even numbers of tube passes may be arranged in series. With identical heat exchangers arranged in series, as shown by Fig. 12(e), the LMTD is computed as LMTD =

A

where

where

T1= shell-8ide inlet temperature T2 = shell-side outlet temperature €1 = tube-side inlet temperature ta = tube-side outlet temperature For the special case that

then equation (7)reduces to LMTD = TI- B, or TI- C

(9) For one-shell and multipl& of two-tube-pass heat exchangers as shown in Fig. 12(b), the.EMTD becomes LMTD = where

X Y

A

Y+X

- T s ) ~+ fh - t~)']"' = TI T2 - (Is + 51) = [(Ti

+

Values for b and m are as follows:

where

lia

m = number of identical single-shell heat exchangers

with an even number of tube passes in series The tube-side film coefficient, hi, may be computed from the Sieder and Tate equation [$I. This equation waa published in 1936 and remains in widespread use; it has the following form:

ho = shell-side film coefficient do = tube outside diameter C, m = constants

Reynolds Number b I

m

0-10 1.4 0.333

10-100 1.36 0.343

100-1000 0.593 0.533

The remaining terms are as defined for equation (12). In the higher Reynolds number range (Re > 1000), There are some variations in the technique8 used in C and m are taken as constants having the values of 0.32 applying equation (13); the variations primarily deal and 0.612, respectively. * ,* with the quantification of C and m. For calculations In a commercial shell-and-tube heat exchanger, there regarding the flow of a gas normal to the tube bank (i.e., is a gap between the tube bundle and shell. A portion a cross-flow heat exchanger), Colburn recommended in of the fluid in the shell will therefore bypass the tube 1033 that equation (13) be applied as follows: bundle, and not be cooled or heated. Also, there is flow leakage between cross-flow baffles and the shell. m = 0.6 It is essential that these non-ideal conditions be Z and Zware assumed equal accounted for, especially in any effort to reduce uncerG = mass velocity of fluid through the minimum flow tainties of flow pressure drop prediction. Tinker [ll) cross-sectional area of the @be tank and Bell [lo] each proposed simplified methods to C -- 0.33 for staggered tube pattern (triangular or correlate these complex flow phenomena to geometry rotated square pitch) factors. Their basic approach is to divide the shell C = 0.26 for in-line tube pattern (square pitch) flow area into three categories, namely, cross-flow area When making computations for the flow of a fluid (the flow area between tubes in the direction normal to through a circular baffled heat exchanger; Donahue the tube axis); leakage area (the flow area between the recommends that 'equation (13) be applied with the tube and tube holes in the cross-flow baffles and the area following assumptions [9]: betweed the b d e outside diameter and the shell inside diameter); and bypass area (the flow area between the m = 0.6 tube bundle and shell). Bell's is an overall approach; G = (G,G,)lln with known relative values of bypass and leakage area G, = mass velocity normal to the tube bundle to that of cross-flow, he, obtained empirically correction Gw = mass velocity through cross baWe window (i.e., factors to be applied tofdeal tubecbank heat transfer baffle cut out area, Fig. 6[a]) coefficient and pressure drop values. Tinker assigned The geometrical mean value of the mass velocity is used flow resistance constants to each flow area and calculated in recognition of the fact that the direction of fluid flow an effective cross-flow rate to be used in the ideal tube hank cross-flow heat transfer and pressure drop correlais not normal to the tubes in the baffle window. On the basis of test data with segmental baffled heat tions. Tinker's is a more useful approach in that refineexchangers, Donhue further recommends that C be ment and generalizations are easier with his method. I t may be seen that computations for the shell-side given the value of 0.22 as an average value. This value

MARINE ENGINEERING

NOTE: Nomograph is based on 518" O.D. #18 BWG t u k For tubes of other O.D., Q end I.D., di, COW ths hio read a follows:

fig. 13

Rlm mo(ficierd for water inside hbw

HEAT EXCHANGERS

NOTE: This nomograph is valid for plain tuber having a 114" O.D. on e 11132"- A pitch or a 3/8"O.D. on a 17/32" A pitch.

MARINE ENGINEERING NOTE: This nomograph Is valld for 518" O.D. plain tubes on a 13/16" A pitch

HEAT EXCHANGERS

NOTE: This nomograph is valid for 518" O.D. t u k with a triangular a d rotated-rquare pitch. . For other tube diameters, multiply chart reedings by do, in.

Multiplier

Fig. 15 Etfective flow area In wgmmtal b d e d flow

Fig. 16 Fltn coefficient for water in segmental b d b d flow over plain tubes

1

504

MARINE ENGINEERING

HEAT EXCHANGERS

II

film coefficient are tedious and time consuming, and a particularly troublesome aspect of the calculation is the assessment of the effective flow area for segmental baffled flow. To simplify calculations, Figs. 14 and 15, which are based on references [ l l ] and [12], were prepared to provide an approximation of the effective flow area. Figures 14 and 15 cover the tube sizes and pitches normally used with marine heat exchangers. It should be noted, however, that manufacturing tolerances on the shell, tubes, and baffles vary and their variation can appreciably affect the value of the effective flow area, especially in the case of small heat exchangers. Therefore, these nomographs, which are based on a particular set of tolerances nominally followed by heat exchanger manufacturers, are of value primarily as an indication of the influence of the various factors which enter into the design of shell-and-tube heat exchangers. .For instance, it may be seen from Fig. 14 that small b d e spacings (small in relation to the shell diameter) are progressively less effective in creating small effective flow areas (or high velocities). These nomographs are particularly useful as devices for checking a design and narrowing down some of the parameters (shell size, tube size, b d e spacing, etc.) in the course of designing a heat exchanger. Figure 16 is a nomograph which relates the shell-side film coefficient to the water temperature and mass velocity. Figure 16 is based on the data presented in reference [lo]. Two means have been commonly employed to allow for the additional thermal resistance when scale or deposit begins to accumulate on the tube wall. The most ri&rous means is to assign values to the tube inside and outside fouling resistances, rdio and rh, as a function of fluid type, temperature, velocity, etc., and compute the required heat transfer surface accordingly. I n applications with mild rates of fouling, however, the fouling resistances are difficult to separate and evaluate meaningfully, and furthermore, they are small. Consequently, the "clean factor" concept has found general acceptance in conventional steam power plant and marine heat exchanger applications. The clean factor is a factor less than unity by which the clean overall heat transfer coefficient, U, is multiplied to allow for fouling. I n order to illustrate the procedure employed when performing heat transfer calculations for a shell-and-tube heat exchanger, consider such a heat exchanger with the following characteristics: Shell size = 6.075 in. I D Tube size = % in. OD 20 BWG (plain tube) Tube material = Admiralty metal Number of tube passes = 2 Number of tubes = 36 U-tubes Tube length = 44 in. Tube pitch = 1 x 2 in. Cross baffle spacing = 2% in. Effective heat transfer surface = 24 sq f t

With 50 gpm of fresh water at 78 F entering the shell side of the heat exchanger, the outlet temperatures and heat exchanged when the fresh water is cooled by 37 gpm of chilled water at 50 F entering the tube side will be computed. First, it is established that the tube inside and outside diameters are 0.305 in. and 0.375 in. respectively. The water velocity through the tubes is computed to be 4.5 fps, and from Fig. 13, the uncorrected film coefficient inside the tubes is found to be 750 Btu/hr-sq ft-deg F. Correcting for the tube size. results.in a film coefficient hi, of 820 Btu/hr-sq ft-deg F. With the stated values for the b d e spacing and shell ID, the effective flow area is determined to be 6.1 sq in. from Fig. 14. The shell flow rate was given as 50 gpm which then corresponds to a mass flow rate, G, of 164 lb/sec-sq ft. Entering Fig. 16 with this value for G and a water temperature of 78 F gives a shell-side film coefficient, h,, of 1735 Btu/hr-sq ft-deg F. The tube metal resistance, i.,, is readily calculated by employing the following equation for heat conduction acrom a cylindrical wall: I, =

NOTE: Nomograph is valid for turbulent flow with a specific pmity of unity within 518" O.D. #18 BWG tuk "

A, log. (doldi) 27rLK

where L is the tube length, K is the metal thermal conductivity, and A, is the tube outside area. For a smooth tube, A, is equal to 7r dJ, and the tube metal resistance reduces to I, =

do log. (do/&) 2K

Since K = 70 Btu/hr-ft-deg F for Admiralty metal, equation (16) gives an r, value of 0.0000461 hr-sq ft-deg F/Btu. Based on experience, a clean factor of 0.9 is considered appropriate for this service; therefore, the scale resistances, r& and rdi,, are considered to be zero. This being the case. substitution of the com~uted values into equation'(4) gives an overall heat transfer coefficient, U, of

+ 0 + 0.0000461 + 0 + &

1

1735

= 544 Btu/hr-sq

ft-deg F

When the clean factor of 0.9 is applied, there results U

=

(0.9)(544)

=

490 Btu/hr-sq ft-deg F

and this is used as the design value. It may be noted that the effect of the term corresponding to the thermal conductivity of the tube wall material is small when compared to the effect of the film resistance; for this reason the thermal conductivity of the tube wall is often neglected. I n order to compute the quantity of heat transferred, equations (6) and (10) are used in conjunction with the following heat balance relationship :

ma. 17 R.roun drop idde tub-

MARINE ENGINEERING

HEAT EXCHANGERS

E = exp

(2:)

The two specific heats are unity in this case, therefore all terms can be evaluated. Tzis found to be 69.8 F and t a is 61 F. With these values known, the heat transferred is computed as:

2.2 Pressure Drop. The film coefficients (and consequently the heat transfer rate) increase with an increase of flow velocities, as noted from the previous discussion; but on the other hand, the flow pressure drops due to friction are proportional to the velocities raised to a power between one and two. Consequently, a design tradeoff must be made which entails striking a compromise between high film coefficients and high pressure drops. The design of a heat exchanger becomes a matter of balancing the saving of heat transfer surface with the cost of pumping power to the exchangers. As a result, a major portion of the design work is expended in evaluating flow pressure drops for various operating conditions and exchanger configurations. In practice, the system designer very often must specify the allowed pressure drops for heat exchangers without an accurate knowledge of their impact on the design of the heat exchangers. There are three courses of action which are frequently taken when it is necessary to specify the pressure drops permitted with heat exchangers: Allow a flat 10 psi pressure drop per stream per heat exchanger. This rule-of-thumb approach is popular in some areas of the industry. The engineer has preferred to stay within the 10 psi limit since higher Fig. 18 Tube diameter and specific gravity correction faclora for tube pressure drop pressure drops with higher velocities may approach the erosion/corrosion limit of the metal and furthermore the resultant high heat transfer coefficient may be excessively sensitive to scaling. The 10 psi maximum allowable pressure drop is also high enough to keep the thermal where designer from "tail-chasing " (i.e., the lower the velocity, W. = shell-side flow rate, lb/hr the lower the heat transfer coefficient; the lower the heat W t = tube-side flow rate, lb/hr transfer coefficient, the larger the heat exchanger; the C , = shell-side fluid specific heat, Btu/lb-deg F larger the heat exchanger, the larger the pressure drop, C p t = tube-side fluid specific heat, Btu/lbdeg F which requires lowering the velocity, etc.). Specify alternative pressure drop limitations. and the fluid temperatures are defined as before' With the wide application of the computer to heat Equations (6), (lo), and (17) can be combined to express the two water outlet temperatures in terms of known exchanger design, this approach does not incur an excessive engineering load in identifying alternatives, and values. The resulting expressions, arranged in the order selecting the optimum alternative can be advantageous. of the computational procedure, are as follows: Use design charts to obtain an approximate drop assessment of t i e heat exchanger size and involved before establishing the design criteria.

Fig. 19 Equlwbnt number of tube rows f w bdffled-fiow prenun drop calculation

MARHE ENGINEERING

HEAT EXCHANGERS

509

4 NOTE:

This nomograph is valid for 518" Wolverine tubes type SIT on a 13118" A pitch.

5 /

8, 7 '\

fig. 20

Pressure drop in segmental baffled flow of water w e r plain tuber

In order to illustrate the use of design charts to obtain approximate, but quick, estimates of the pressure drop through heat exchangers, the example problem in Section 2.1 will be continued. Entering Fig. 17 with a water velocity of 4.5 fps and temperature of 50 F gives a pressure drop within the tubes of 0.085 psi/ft. Since there are two tube passes, the total tube length is twice 44 in. or 7.34 ft, and the total uncorrected pressure drop is 0.624 psi. The inside diameter of the tubes is actually 0.305 in. and the specific gravity of the fresh water is unity; therefore, the corresponding correction factors to be applied to the pressure drop taken from Fig. 17 are read from Fig. 18 aa ad = 1.95 hnd a. = 1. Applying these to the uncorrected pressure drop of 0.624 psi gives a corrected pressure drop of 1.22 psi. In order to calculate the baffled-flow pressure drop on the shell side, the number of baffles must be established. The number of tube baffles can be estimated as tube length -1 = 44 -1 = 18 2.25 baffle spacing I n addition, the number of tube rows which the flow crosses between each pair of baffles and the baffle cut-out area (baffle window) must be established before the calculations can proceed. It is convenient to introduce the term N,, which is defined as the equivalent number NB =

of tube rows that the flow crosses between each pair of baffles, as the ensuing calculations are greatly simplified. The equivalent number of tubes citn be determined if details regarding the baffle window height, baffle spacing, tube spacing, and shell are known; but such is seldom the caae in preliminary design work. For estimation purposes, Fig. 19 has been prepared oh the basis of a commonly used heat exchanger configuration to give guidance in determining the equivalent number of tube rows crossed between baffles. Entering Fig. 19 with a shell I D of 6.075 in., a baffle spacing of 2.25 in., and a triangular tube pitch of 17dZ in. gives an equivalent number of tube rows of 11.5. Consequently, the equivalent total number of tube rows is computed as NR = (NB

+ 1)N. = 21.9 rows

The pressure drop per row of tubes can be evaluated from Fig. 20. A line is drawn between a water temper* ture of 78 F and a mass velocity of 164 lb/sec-sq ft (line 1)) and the interception with the reference line is noted. Next, line 2 is drawn from the reference line intercept and the second mass velocity curve, and the product of the fluid specific gravity and pressure drop per tube row of 0.024 psi/row is obtained. Since the fluid specific gravity is unity, the shell-side baffled-flow pressure drop is (0.024) (21.9) = 5.26 psi.

Fig. 21

Meciive flow area for segmental baffle flew

MARINE ENGINEERING

HUT EXCHANGERS

51 1

that the resistance to heat transfer increases rapidly initially, but the rate of increase subsequently drops to a very sm* value. The tests h indicated that the fouling characteristics of finned tubes are similar to those of plain tubes, and both types of tubes have comparable percentages of reduction in performance. At the end of the four-week test period, ,the fouling resistances for both fin-tube and plain-tube bundles were below the value of 0.005. During the tests, both the finned and plain tubes were cleaned by a kerosene-water-detergent emulsion and the heat transfer rate waa restored to that initially achieved. In order to illustrate the princjples involved in the e,pplication of finned-tube heat exchangers, consider such a unit with the following characteristics: Shell imide diameter, Di = 23.265 in. -- Tube size = %-in. low fin Tube material = aluminum Number of tube passes = 2 Number of tubes, N = 652 Tube length, L = 96 in. in. Tube pitch = B&e spacing = 9 1 % ~in. Tube length between tube sheets, L, = 89% in.

h = ,'

B~UIM~Z-F

1

% fig. 23

A check will be made to determine the suitability of this

+

Rdo

Fin resistance of low-fin tubes

unit for the following service: Shell Side

NOTE: Multiply ho mad above by (2/z1,-~)-14

when, Z = viscosiw at average tempenturn, centipoi~er & 5 vMcosiw at tube wall temperature, centipdnes

fie. 22

t

1

Rlm coeffidenf ftw oil in segmental baffled tlow wer low-tln tubes

2.3 Design of Fin-Tube Heat Exchangers. Another form of the'basic heat transfer relationship given as equation (1) can be written as

where Ra, and Ro are the scale resistance on the inside and outside of the tubes respectively, Rwis the resistance of the tube, and the remaining symbols are as previously defined. If the shell-side film coefficient, ho, is small (as is the case with gases and oils which have low thermal conductivities) in comparison with the tube-side film coefficient, hc the shell-side resistance, (h,A,)-1 will control or "bottleneck" the heat flow. A means of counteracting this circumstance is to increase the tube

outside surface area, A,, to the extent that hoAo= hiAi, in which case the heat flow would no longer be choked by the shell-side heat resistance. The heat transfer surface on the outside of the tubes can be effectivelyincreased by providing fins on the tube outside surface. A type of finned tube frequently used in shell-and-tube heat exchangers is the so-called "low-fin" tubes which have a ratio of finned surface to tube outside surface in the range of 3 to 4. In applications where the tube outside film coefficient is as small as 50 percent of the tube inside film coefficient, the choice of low-fin tubes will permit a more economical, compact, and light unit than a similar design employing bare tubes. Regarding the fouling of finned tubes, tests [13] were conducted with 160 F, No. 5 fuel oil flowing through the shell side of a heat exchanger. The test results showed

Fluid. .................. Flow rate, W, lb/hr.. .... Specific gravity at average temperature. . . Thermal conductivity, Btu/hr-fMeg F ....... Viscosity: 2,centipoises at 210 F.. ............ at 167 F.. ............ Specific heat, C, Btu/lb-deg F. ......... Inlet temperature, F. .... Outlet temperature, F . ...

oil 123,400 0.84 0.0665

106 280

Tube Side water 66,500

pute the tube-side film coefficient. The tube inside diameter is 0.402 in., the average fluid temperature is 158 F, and the tube-side ve10lcit"~ is computed as 1.03 fps. A value for h, of 435 Btu/hr-sq ftaeg F is taken from Fig. 13 for Yrin. OD 18 BWG plain tubes which have an outside/inside area ratio of 1.186. Low-fin tubes have an inside/outside ratio of 0.26; and therefore the correction for low-fin tubes is made as follows: hio =

( L ~ ) Ai

plain tube

(4) Ao

0.527

O''

finned tube

= 141 ~tu/hr-sqft-deg F

To compute the shell-side film coefficient, the effective flow area must be known; this is determined to be 146 sq. in. from Fig. 21 by entering with the stated values The quantity of heat to be transferred is computed as for the shell I D and the baffle spacing. The effective W,C,(T1,- Tt) = 1,720,000 Btu/hr, and the LMTD is flow area, in conjunction with the shell flow rate, gives determined to be 47.3 F from equation (10). an effective mass velocity of 33.8 lb/sec-sq ft. The Low-fin tubes having an outside diameter of % in. shell-side film coefficient can be determined from Fig. 22 have a finned surface area of 0.405 sq ft per linear foot; by drawing a line (line 1) between the viscosity and mass therefore, the total finned surface, A,, is 0.405 NL, = velocity scales and noting the intercept on the reference 1970 sq ft. line. Next, a line is drawn from the thermal conducI t is now possible to compute the required overall heat tivity scale through the ina&cepf poia on the rehrence transfer coefficient from the expression line, and the intercept with the h. scale is read as 22.5 Btu/hr-sq fMeg F. This value must be corrected, howU = Ao(LMTD) = 18.45 Btu/hr-sq ft-deg F ever, to account for the variation in viscosity at the tube wall. Since the temperature of the oil at the tube wall If in this service the unit will provide an overall heat is not known at this point in the calculation, it is necestransfer coefficient of this magnitude, it would be sary to assume a temperature and then confirm the considered satisfactory. The h t step in obtaining an assumption to be satisfactory when it is subsequently estimate of the heat transfer rate for the unit is to com- established. With an assumed tube wall temperature

i

I

I

11/ II

I

512

The assumed tube wall temperature upon which the calculation was baaed must now be checked. The calculated tube average wall temperature is 1, =

Tmg-

(LMTD) = 168.7 F

-

This is reasonably close to the assumed temperature of 167 F ; therefore, the viscosity correction factor of 0.873 is sufficiently accurate for practical purposes. The calculated clean overall heat transfer coefficient gf 17.2 is somewhat smaller than the required overall heat transfer coefficient value of 18.45; consequently, no allowance for fouling during service is provided. In an actual design, the heat exchanger design should be modified to improve the unit's heat transfer characteristics. This could be accomplished by decreasing the baffle spacing (so as to increase the flow velocity and increaae the heat transfer coefficient) if the associated increase in pressure drop can be accommodated; or an alternative solution would be to increase the amount of heat transfer surface. Tests which have been conducted to corroborate the calculation procedure outlined in the foregoing have shown the method to be valid. Measured heat transfer coefficients have been slightly greater than those calculated, which is a desired characteristic for the difference to possess. The fundamental relationship for computing the pressure drop on the shell side of a finned-tube heat exchanger has the form

this line, use Fig. 25 for turbulent flow

.2J

ffg. 24

5 13

HEAT EXCHANGERS

MARINE ENGINEERING

I

Pressure drop In wgmontal baffled Aow wer low-fln tuber

of 167 F, the corresponding viscosity is 280 centipoises. 19.7 (aasurning that the shell-side fouling resistance, R*, This, in conjunction with the average oil temperature of is zero) the resistance of aluminum h, rti., is seen to be 210 F and viscosity of 106 centipoises, gives a viscosity 0.00013 hr-sq ftdeg F/Btu. As may have been anticicorrection factor of 0.873 and a corrected shell-side hlm pated, the resistance of low-fin tubes is quite small for their usual applications. coefficientof 19.7 Btu/hr-sq ftdeg F. The thermal resistance, r., of the annular ring portion Another factor, in addition to those involved with the analysis of plain tubes, is the thermal resistance of the of the fin tube wall is not included in the fin resistance fins themselves. When computing 'the heat transfer taken from Fig. 23. However, it is readily computed coefficients of finned tubes, an allowance must be made from equation (15), and is found to be 0.00012 hr-sq ftfor the fins because, aa part of the tube walls, they offer deg F/Btu. The estimated overall heat transfer coefficient can now some resistance. An estimate of fin resistance may be made from Fig. be evaluated aa 23, which presents the fin resistance of low-fin tubes of 1 diierent ,materials in common use. The fin efficiency, = 1 = 17.2 Btu/hr-sq ftdeg F upon which the resistances shown were baaed, was taken h, rti. r, from reference [14]. With a shell-side film coefficient of

+ + +

NOTE:

1.000

nomograph is for E/S" low-fin tuber on a 13/16" A pitch and Reynolds numben greater than 1800.

800 600 400 300

a2 /

,' I''

0.0003

where sg is the specific gravity of the shell-side fluid, and the other terms are aa previously defined. The viscosity term is known aa it was evaluated in connection with the determination of the shell-side a m coefficient. The number of baffles is given aa 8 and the fluid specific gravity is 0.84. With the shell and t@e data previously given, the equivalent number of tube rows crossed per . b a e d space is found to be 27 from Fig. 19. Figure 19 is applicable to both low-fin and plain-tube configurcc Fig. 25 Pressure drop in segmental baffled flow wer low-fln tuber (turbu1e.d tions. Note also that the effective flow a r e a evaluated flow range) from Figs. 14, 15, and 21 are used for both beat transfer and pressure drop calculations; this is a result of the approach taken in correlating the test data. velocity pcale, and the intersection with reference line B Only the (AP X sg) term remains to be determined. is noted and defined aa reference point 2. Finally, A value for this term, can be obtained from Fig. 24 or line 4 is drawn through reference point 1 (line 2 and Fig. 25. Entering Fig. 24 with the previously established viscosity scale) and reference point 2 (line 3 and reference viscosity of 106 centipoises and mass velocity of 33.8 line B), and at the intersection with the sg X A P scale, lb/sec-sq ft, line 1 is drawn between the viscosity scale the value 0.02 psi/row of;Abes ia read. and the left-hand mass velocity scale. At the point line With all terms in equation (19) evaluated, the shell1 crosses reference line A, line 2 is drawn along and in side pressure drop is found to be 6.64 psi. between the inclined solid guidelines which are provided The tube-side pressure drop is calculated the same aa to ensure that line 2 has the proper slope as it intersects with plain tubes. With an average water temperature the viscosity scale; note the intersection with the of 102.5 F and a water velocity of 2.55 fps, the pressure viscosity scale and define it aa reference point 1 (it haa drop per foot of tube length is 0.026 psi/ft, from nothing to do with viscosity). Next, line 3 is drawn Fig. 17. The tube diameter correction factor, a d , is 1.4 between the viscosity scale and the right-hand mass and the specific gravity correction factor, a,, is unity.

514

HEAT EXCHANGERS

MARINE ENGINEERING

Therefore, with an 8-ft tube length and two passes, the tube-side pressure drop becomes The method employed to formulate the nomograph for rrssessing the heat transfer and pressure drop within heat exchangers is bersically that of Tinker [ll], Devore

[12], and Donahue 191. Details of the derivations and assumptions involved with the method have not been reviewed as they are not of primary interest to the ~racticine:engineer. The Drocess of mine: through the procedure ouiined will, hoGever, convey an appreciation of the oonsiderations involved in the hydrodynamic and thermal design of a shell-and-tube heat exchanger.

- -

P.T. WNN-NEAR & FAR SIDES

a

515 P.T. VENT-NEAR SIDE PLUGGED /

OIL W N N .

\

PIPE TAP (P.1.)

Section 3 Heat Exchanger Applications 3.1 Lubricating-Oil Coolers. In addition to serving as a lubricant between moving mechanical parts, lubricating oils generally also accomplish a second objective of removing the frictional heat generated. Therefore, some means must be provided for removing the heat absorbed by the lubricating oil. With small systems, the natural heat transfer by radiation and convection may be adequate; but with larger systems, particularly those employing forced circulation, lubricating-oil coolers are required. The main lube-oil coolers (i.e., those in the lube-oil system serving the main turbines, reduction gears, and main thrust bearing) are generally the largest used aboard ship. A shell and straight-tube exchanger with a removable tube bundle and a double-packed floating tube sheet type of construction (such as illustrated by Fig. 26) is the design most commonly used for main lube-oil coolers. Oil flows in a single pass via transverse bafaing in the shell, and seawater is the normal coolant flowing in one or more passes through the tubes. The floating tube sheet is centered between the shell flange and waterbox flange. A gland ring retains the packing. Separate packing rings are provided for the shell (lube oil) side and the coolant side. The gland ring is grooved around the inside, and leak-off holes are provided so that leakage past the packing rings on either side will be relieved to the outside and attract the operator's attention. The stationary tube sheet, baffles, and support plates are assembled and held in proper relative position by tie rods and spacer sleeves. Tie rods are threaded into but not through the stationary tube sheet. The waterbox and stationary tube sheet are secured by collar bolts or by stud bolts driven into tapped holes in the tube sheet so that the tube-sheet to shell-flange joint will not be broken when the waterbox is removed. The packing retainer ring is usually "scalloped" so that there is a stud hole for every second stud in the shell flange st the floating tube sheet end of the cooler. This enables the gland ring to double as a test ring and allows the shell to be hydrostatically tested without the waterbox in place. In some of the more demanding applications, such as submarine heat exchangers cooled by seawater, there are special requirements for bolting as well as for materials,

and such requirements must be considered in the cooler design. Limitations are also provided for maximum fluid velocities and pressure drop, and shock resistance requirements may also be specified for military applications [6]. In the late 1950's, it was determined that the performance, in terms of fouling properties and service life, of low-fin tubing is approximately equal to that of bare tubes, and a gradual change has taken place such that %-in. and %-in. low-fin 90/10 CuNi tubing has become the preferred tubing in lieu of the previous %-in. and %-in. bare tubes. With low-fin tubes, the greater amount of heat transfer surface within a given shell size outweighs the disadvantage of the somewhat lower heat transfer rates of finned tubes, and the net result is a smaller, more compact, and more economical unit for a given performance requirement as compared with bare tubes. Lubricating-oil coolers are also used aboard ship for auxiliaries such as main feed pumps and air compressors. These units are similar in design to main lubricating-oil coolers but are smaller, having a shell diameter of 6 to 10 inches. The thermal design of oil coolers is treated in detail in Section 2. 3.2 Fuel-Oil Heaters and Lubricating-Oil Purifler Heaters. The residual fuel oils usually burned aboard ship are so viscous at atmospheric temperatures that they must be heated before they can be pumped from the storage tanks (the properties of fuel oils are discussed in detail in Chapter 23). The more viscous fuels may require heating to 120 F or more before they can be pumped; and when supplied to the burners, the fuel oil must be at an even higher temperature in order to attain a fuel viscosity sufficiently low for proper fuel atomizrlr tion in the burners. Therefore, two stages of fuel-oil heating are provided; the primary heaters are installed in the fuel-oil tanks, and the secondary heaters are installed between the service pumps and the burners. The primary oil heaters are installed in the tanks in either of two forms: as steam-supplied pipe grids or coils or as open-ended tank-suction heaters. The latter alternative has become increasingly popular due to (a) lower initial cost, (b) lower maintenance costs, and (c) lower steam consumption (only the oil to be pumped is

\ SFE DETAIL "A"

LEGEND, 1. Shell 2. Water Channel 3. Channel Cover 4. Return Channel 5. Fixed Tube Sheet 6. Flpeting Tube Sheet 7. Tubes 8. Baffle Assembly 9. Retainer Ring 10. Packing 11. Shell Flange Gasket 12. Channel Gasket 13. W r r Gasket 14. Zinc Amdo# (wha~Imludd)

SHELL FLANGE

GI

DETAl L "A" Detail of Floating Tube Sheet End Section at pecking retaining r i ~ g w e p h o l e .Mixing of the shell and tube side fluids thru the packing is impossible with this detail. Any seepage from either side, rwulting from loosening of the bda, d r i p out thru the copper-lined packing ring weep-holes, thus putting the m t i n g personnel on notice. The nuts on both sides may be made up to tighten packing without hutdown or interruption of operation. fig. 26

mg. 27

Typical main lube-ail c o o k

Typicol tank rudon fuel-oil heator

MARINE ENGINEERING

heated as opposed to heating the entire tank). A typical tank-suction fuel-oil heater is shown in fig. 27. hi horizontal U-tube heat exchanger has an outer shell flange that is bolted directly to the tank. The oil is drawn in through the open end of the shell and across the tube bundle. As the heating medium (condensing steam) circulates through the tubes, the portion of oil in the heater shell in contact with the tubes rises in temperature with a corresponding decrease in viscosity. The entire process is continuous with the oil flow being

HEAT EXCHANGERS

CLEANING FLUID INLET

MOUNTING BOLTS

RELIEF VALVE CONN.

from those surfaces. Steam is delivered to the steam chamber, and from there enters a series of small tubes, cdled "inner" tubes, through which it travels until it is dpcharged into the annulus between the inner and outer tubes at the far end of the outer or oil-heating tubes [see Fig. 8(b)l. Thus steam (free of condensate) is in contact with the heating surface at that part of the surface where the highest oil temperature is desired. The space between the inner and outer tubes is small so that the volume of steam flowing in the annular space generates a velocity of flow sufficient to continuously sweep the condensate from the surface as it forms, thus reducing water-film losses and preserving high transfer rates on the steam side.

CLEANING FLUID OUTLET

,

PRESSURE GAGE CONN.

serves aa a shell. The Navy haa set forth specific design criteria for all three of the foregoing designs, and further describes an "evaporator" type of fuel42 heater which uses an intermediate fluid to transfer heat from a primheating coil to the bayonet oil-heating elerpents; this precludes any possibility of contaminating the condensate with fuel oil in the event of a leaky tube or tube sheet joint [15]. Steel construction throughout is standard practice. A minimum of two heaters is normally provided, sized and so arranged that either heater can provide full boiler requirements at overload with the other heater serving as a standby. Heaters are fitted with a relief valve which usually discharges to the settling tanks via a check valve. The automatic temperature control valve is normally installed in the steam supply line with the sensing element located in the oil outlet flow line immediately adjacent to the heater. The steam flow valve is usually installed immediately adjacent to the steam inlet connection to the heater. Adequate steam traps are necessary in the condensate lines from the heater since flooding of the heater will have an advem effect on its performance. Sectionalized heaters or banks of heaters are commonEy used in order to provide the flexibility of using all or a portion of the heating surface over a wide range of heating capacity (from minimum port steaming conditions to 120-percent ovedoad). With such an arrangement, individual heaters or sections of a heater can be cut in or out as the demand fluctuates and thereby maintain a steam supply suflicient to ensure adequate control of the oil outlet temperature. Such a control of the amount of heat transfer surface in service avoids cyclical heater operation wheh the heating requirements are very small as compared with the effective heating

517 I

I

I

I

SECONDARY

The design requirements for lubricating-oil purifier heaters are much like those for fuel-oil heaters in that the oil n~ustbe heated to a prescribed twperature range (normally 100 to 160 F) in order to attain a sufficiently low oil viscosity for effectivepurification. Lubricatingoil purifier heaters are generally of the tubular type and are similar to those used as fuel-oil heaters.

I

3-3 Boiler Feedwater and Desuperheater Leakage Test Sample Coolers. Daily tests of the condition of the

fig. 28

Bayonet-tube fuel-oil heater

boiler water and feedwater are necessary to ensure continued efficient operation and protection of highPressure steam generators. These tests are necessary to : Maintain the specified boiler water chemistry through chemical treatment to ensure that the correct proportions of the essential chemicals are present. Check the effectivenessof the blowdown procedure by the concentration of the soluble and suspended solids in the boiler water. Determine the amount of dissolved oxygen in the boiler water and feedwater to guard against excessive The collection and cooling of water samples is the first in the test procedure. The coolers required

mi. 29

Boiler water sample coder

are relatively s m d heat exchangers due to the small quantity of sample requifd for testhg. One of the common arrangements is a cooler system consisting of primary and secondary coolers connected in series. The coolers have a cylindrical cast bronze shell that contains a helical mil wound around a core positioned in the c a t shell; such a cooler is shown in Fig. 29. The sample p=es through the coil and cooling water flowa across the outer surfaces of the coil. The boder water sample cooler coil is usudy made, of 90/10 or 70/30 CuNi and

MARINE ENGINEERING

HEAT EXCHANGERS

66

519

Fig. 30 Saltwater heater and drain cooler

the cooler and valves are designed to the boiler working cooler uses seawater at 85 F as pressme. The a and reduces the sample temperature to cooler uses chilled approximately 140 F; the 50 F fresh water as the coolant to reduce the temperature of the sample to that desired for testing. The maximum temperature for oxygen determination is 70 F; in Navy practice100 F is considered the maximum temperature euitable for pH, hardness, and chloride determination. The coils far desuperheater leakage test sample coolers can be madeof copper or a similar material since they me designed for a'rnoderate pressure and are normally cooled with fresh water. 3.4

Tank

Cleaning System

Heaters

and

Drain

Coolerr. When a different grade of oil is to be carried in a tanker (e.g., gasoline vice crude oil), it is necessary to wash the cargo oil tanks before receiving the different grade. The tank cleaning systems use hot seawater supplied from heat exchangers. Cargo vessels utilize a smaller system than tankers for tank cleamgThe heat, exchangers usually consist of a seawater heater rwlddrain cooler connected in series SO that the seawater flows first through the drain cooler and then through the heater. Steam is used as the heating medium in the heater, and the resulting condensate is cooled in the drain cooler. The seawater is heated to a temperature of 180 to 200 F. As illustrated in Fig.30, the exchangers are typically of the horizontal shell-and-tube type; the heater (above the d r h cooler) is of the U-tube type, and the drain cooler has a fixed tube nest and a shell expan~onjoint. The tube nest in the coolers is usually fitted with transverse baffles to create a flow path for the drains perpendicular to the run of the tubes. Depending upon design

conditions and installation requirements, the tubes may be arranged for single- or m ~ l t i - ~ a flow s s of the seawater. Zinc anodes are provided in the water heads to minimize galvanic corrosion on other parts of the heat exchanger (galvanic corrosion is discussed in section 1.1). Occasionally the heater and drain cooler are combined preclude the in a single shell, but the arrangement possibility of the condensate rising above its normal level and submerging the tubes in the heating section, which would reduce the effectiveness of the unit. In some cases, only a heater is installed and the drains are discharged through a steam trap to the deaerating feed heater or a suitable vessel or receiver. Where a separate or combined heater and cooler is installed, a liquid level control is employed to ensure submergence of the drain cooling tubes. Relief valves are provided on both the shell and tube sides of the saltwater heater. For naval applications, there are a number of special design requirements (such as construction materials and shock requirements) which must also be considered [16I. 3.5 LOW-PressureFeedwater Heaters. h he classification (i.e., low pressure ar high ~ressure)of feedwater heaters depends upon their location relative to the boiler feedwater pump; low-pressure heaters are located on the suction side of the main feed pump, whereas highpressure heaters are located on the discharge side. Feedwater heating is accomplished in a number of steps or stages and the heaters are usually referred to as the first stage, second stage, third stage, fourth stage, etc. Multiple stages of feed heating are essential to the efficiency of a steam turbine power ~ l a n t ,as may be t balance estabnoted from Chapter 2. The ~ l a n heat lishes the number of heating stages, feed flow through

I each point and a k l i a r exhaust ~ pressures, furnished with the glsnd-exhauster fan and and the temperature of the feedwater entering each drain r e d a t oa rpackage mounted on the unit. Due to the of Since the heat transfer coefficient of condensing steam combined heattemperature exchangers,difference the outletbetween ends ofsections the tubes

is

of velocity, and feed pressures are usually quite high, the feed is generally in the tubes with the in the For a given steam Pressure, the heat is dependent upon the feed velocity through the tubes. of to 7 fps result in a reasonable pressure and satisfacto~heat h n s f e r conditions. The heater should avoid the Occurrence of dead spaces, and drain cooling sections should hold 'lose baffle-to-shrOud tolerances so as to avoid excessive bypasing beat transfer surface which would result in insdequate drain caoling. Both the shell and water sides should be self-venting. The pre-e CQntr0l.i the constlTl~ti0ndetails feedwater heaters with design pressures UP to PS~Fare considered low-pressure heaters. It is common practice to combine Several lowpressure heaters one shell to save space, C O S ~of equipment, piping, and installation costs. Figure 31 depicts a ty~lcalcombined low-pra4sure feed h e a t e d r a i n cOO1er/dand-exhaust condenser. similar heaters combing two Stages of h e a t h with the drain cooler are sometimes used. The heater is normally

me u s u d y secured in the tube sheet by of alternate rings of metalic and fibre packing (see 1,-ig. 8 ( ~ ) and ) are therefore free to expand independently. 3.6

Dim.Montact

Dwerating Feedwabr ~

~

Since marine boilem are operated at high temperatures and pressures, there is a hazard of attack due to the presence of dissolved oxygen or carban diofide in the feedwater. It is ~ t u a l l y impossible to prevent the entry of air into the feed system, particularly during plant st&tup; therefore it is necessary to provide deaerating equipment for the removal of air and co~osive gases from the boiler feedwater. Although &aeration can be largely acomplbhed in the condenser, ucondenser deaeration' is not suffci$n&during @ant stertup; and without further &aeration there be no-provision for the removal of air introduced later in the system, particularly a t the condensate pumps. Flash &aeration, whereby saturated water at 10 to 15 psig is introduced into a surge tank at atmospheric pressure, is simple and economical. H ~ the ~ resulting flashing of steam will not ensure the low dissolved oxygen content (i.e., 0.005 cc per liter) reqUired

~

HEAT EXCHANGERS

MARINE ENGINEERING

VENTED STEAM CONTAINING OXYGEN REMOVED FROM WATER

WATER SUPPLY WATER SPRAY

VENT CONDENSING WATER SPRAY NOZZLE PRIMARY HEATING

AND DEAERATION CHAMBER

SPRING LOADED ATOMIZING VALVE

Fig. 32

Direct-contactfeed heater

of the feedwater unless suflicient agitation is positively provided. There is also a loss of steam through flash deaeration. For example, a deaerator operating at 15 psig flashing down to atmospheric pressure loses about 4 percent as flashed steam. The flashed steam should not be condensed and returned to the feed cycle since it will normally have re-entrained a portion of the undesirable dissolved gases. Therefore, standard marine deaerators are more sophisticated in design than those of the flash type. The practical considerations involved in the removal of dissolved oxygen from boiler feedwater may be briefly summarized as: 1 Heating the water to the boiling temperature for the pressure under which the process is conducted (satura tion conditions). From the chemical relationship termed "Henry's Law," it is known that when a partial pressure of a liquid is equal to the total pressure above the liquid

(boiling conditions), the solubility of any gases in the liquid is zero. 2 Providing a design that ensures thorough agitation and scrubbing of the feedwater by the steam. Complete agitation of the feedwater and contact with the scrubbing steam ensures that equilibrium will be reached and that the zero potential solubility condition (Henry's Law) will be attained. 3 Continuously venting from the system a mixture of gases and steam. Through the use of adequate venting, the partial pressure of the noncondensable gases in the system will be kept low and the saturation boiling point of the liquid will be maintained. The heater immediately preceding the suction side of the boiler feed pump is usually the "direct contact" or deaerating feed heater (generally known as a "DFT" for deaerating feed tank). A typical direct-contact feed heater is illustrated in Fig. 32. Condensate and

makeup are sprayed into the steam-filled primary heating and deaeration chamber through a series of spray nozzles and a vent-condensing spray nozzle. The spray nozzles provide an even distribution of water over the entire heating area. The steam flow, which- is essentially counter-current to the water flow, heats the water close to the saturation temperature such that the solubility of the gases is zero, and approximately 95 percent of the oxygen content is thereby released. Water and condensate collect in the conical water collector and flow to the atomizing valve, where highvelocity steam strikes the mixture, atomizes it into a fine mist, and raises the temperature the last few degrees to its saturation point. The mixture strikes a deflecting baffle which separates the water and steam. The hot gas-free water drops to the storage compartment. The complete atomization and heating of the feedwater by the steam jet ensures that the dissolved gases will be released. After the atomization process, the steam and released gases flow through the primary heating/deaerating chamber where a large portion of the steam is condensed as it heats the incoming water. A small portion of the steam and all of the gases pass through the integral vent condenser which condenses the majority of the remaining steam. The small amount of steam vapor that is mixed with the released gases is then discharged to the atmosphere or to the gland leak-off condenser. The deaerator conditions feedwater such that its dissolved oxygen content is less than 0.005 cc per liter. In addition, it substantially reduces the carbon dioxide content of the feedwater. Since the feedwater is at saturation temperature and above atmospheric pressure, the arrangement of the deaerator is of great importance as there is a strong possibility of the feedwater flashing into steam at the pump suction. There are two means of ensuring an adequate suction head at the main feed pump. One is to position the deaerator high in the maqhinery space; or an alternative is to provide a booster pump between the deaerator and feed pump which will maintain an adequate suction head on the feed pump. Damage control considerations dictate that the booster pump arrangement be used in naval ships. The booster pump must be designed to handle condensate at saturation temperature, and it is important that the booster pump suction line be short with little or no turns and adequately vented so that pump cavitation and suction line flashing will not occur. The alternative of locating the deaerator high in the machinery space is the preferred arrangement with merchant ships as a pump is eliminated and damage control is not a design criterion. Proper performance of a deaerator requires correct sizing of components and control of the rate of flow to the storage tank portion of the deaerator. The first major consideration is the boiler steam output. This determines the size of the deaerator and affects the storage tank, the makeup valve, the transfer pump, and the number of water spray nozzles in the unit. The other

52 1

major factor is the temperature of the water delivered to the spray nozzle; this temperature determines the size of the steam-regulating valve which admits steam to the deaerator. This valve is sized as closely as possible to furnish the quantity of steam required to maintain the deaerator a t the operating temperature, plus about 10 percent additional capacity/of steam over that required to heat the inlet water at the design conditions as a safety margin to handle surges of incoming feedwater. However, sincpc the steam capacity is considerably affected by the pipe size of the regulator and the incoming steam pressure, it is difficult to provide a valve that exactly matches the desired capacity. Proper deaera tion requires that the temperature.of the incoming water be raised to the saturation point; therefore, the volume of the inlet water must be controlled in relation to its temperature to stay within the heating capacity of the steam supplied by the steam-regulating valve. An excessive flow of cool water will, of course, quickly condense the steam in the deaerator, making it difficult to maintain the desired pressure. This emphasizes the necessity to provide an adequate safety mapgin in sizing the steam-regulating valve so that its capacity and response rate are capable of handling surges of cool water. The storage tank is usually selected to retain about five minutes of storage. If high-pressure (high-temperature) returns are available, they may be returned directly to the deaerator storage tank. Here they will flash and provide a certain amount of steam for preheating the water introduced into the dyapator. If these returns exceed 25-30 percent of the total capacity of the deaerator, more steam will be available than is needed and some other means must be employed to use the returns. Deaerating feed heaters are normally equipped with two spring-loaded relief valves: one to prevent a high pressure from accidentally building up within the tank; and the second, known as a vacuum breaker, to prevent a high vacuum from developing in the tank by allowing atmospheric air to enter the tank in the event that the pressure in the tank drops below a prescribed value. The shell and majority of the internals are normally of welded-steel construction; however, the 'steam baffles, spray nozzles, atomizing valve, and vent condenser are generally manufactured from nonferrous alloys or stainless steel. , 3.7 High-Pressure Feedwater Heaters. A high-pressure feedwater heater may consist of one, two, or three sections (a three-section heater contains desuperheating, condensing, and condensate cooling sections). All sections are normally integrkkd in oneshell fpr conipactness and simplicity of piping. In addition to the marine regulatory body design requirements [3, 4, 51, the code requirements of the Feed Water Heater Manufacturer's Association [18] are often applied. The construction features of a typical high-pressure feedwater heater are shown in Fig. 33. The tubes are usually N-in.-OD tubes, arranged on a 1x6-in. triangular pitch. The tubes in the condensing

I

I

1 I

I I

I I

I

~

I

I I

I

I

~ II

1

I

522

MARINE ENGINEERING

1

I (a) HORIZONTAL INSTALLATION

1. CHANNEL 2. CHANNEL COVER 3. STATlONdWY TUBE SHEET 4. TUBES 6. TUBE SUPPORT 6.SHELL SKkRT 7. PARTmKlN PLATE IL PARTITLON COVER S. TRANSVERSE BAFFLES 10. tMPINGEMENl PLATES 11. DESUPERHEATIMG ZONE 12 SUBGOOLING ZONE 13. SHELL 14. HEATER SUPPORTS 15. FEEDWATER INLET 16. FEEDWATER OUTLET 17. DRIP INLET la STEAM lNLET r a CONOENSATEOWLET 20. WELL RELIEF VALVE CONNEmlm 21- TtlEE VALVE CONNEmlON - SIDE -.- - -RELIEF .- a.LIWLD LEVEL CONTROL CONwmlONS % *G GLASS CONNECTIONS 24. OPERATING AIR VENT Q)NNEmIONS

- - - --

(M VERTICAL INSCALLATIOM

Fig.

33 Typical fedwafer haatam

section &t supported by plates spaaed at intepvak n d exmetkg 48 in. to avoid tubs vibration Beetions me cut out of the support plates to pmvide passages for steam flow and drainage. If the steam which enters the shell side of the feedwater heater is hi&ly superheated (e.g., a superheat of 100 deg F or above), the tube amface in contact with the superheated steam will have a wall temperhigher than the saturated steam temperature; this lneans that the tube wall will not be wetted by condensate and thst the transfer of heat will be low d e s s special preeauticr11~are made. It is found economical sometimes essential to ban a desuperheatbg section to mntml the desugerhesting of the steam. The desuperhestiing section is h t e d at the feedwater exit end so that the leaving feedwater can be heated to

the highest possible temperature. By arranging tho

Row this way, the feedwater temperature may eve11 exceed the &teamsaturation temperature in the desuperheating section. The desuperheating section consists of a shroud wrapped around a group of the tubes so as to confine the inlet steam. Cmss baffles are provided with the shroud to decrease the dry-vapor thermal resistance. Other design features incorporated in the desuperheating section are means to shield the other regions of the unit from the high-temperature steam and means to prevent distortion due to unequal temperature distribution. The eonde~mate-coolingsection is Located at the feedwater inlet end of the beater so that the condensate (or drains) from the condensing section can be subcooled to approach the feedwater inlet temperature. The cotl-

i

523

HEAT EXCHANGERS

densate cooling section also consists of a shroud enclosing a portion of the tubes and cross-baffle plates. When designing the condensate cooling section of the unit, the possibility of condensate reheating must be considered. Condensate reheating is caused because the steam condensate inside the shroud, while being cooled by the feedwater inside the tubes, is dso heated by the steam condensing on the outside of the shroud. This reheating of the condensate is a matter of great importance at the drain outlet end of the heater where the temperature difference between the condensate and feedwater is often as low as 10 degrees, while the difference between the condensate and steam outside the shroud can be as high as 100 degrees. Several means can be taken to avoid excessive reheating of the steam condensate. One would be to increase theratio of the condensate cooling tube surface area to the shroud area; another would be to insulate the shrouds. The attitude of the feedwater heater as it will be arranged aboard ship must be established before the thermal design of the feedwater heater can commence because the various alternative arrangements impose different restrictions on the thermal design of the unit. When the unit is installed in a vertical position with the feedwater entrance and exit channel (or waterbox) on the top, as in Fig. 33(b), the bottom region of the shell can be used as a steam condensate collector. The shrouded condensate cooling section extends the full length to the top of the shell in this instance, and the height 'of the unit is relatively short. A vertical arrangement with the feedwater entrance and exit channel at the bottom [the inverse of that shown in Fig. 33(b)] is normally selected for long units that are designed for outdoor land installations; however, a lack of space usually precludes its application in marine plants. With this type of arrangement, the unit is designed such that the steam eondensate exitcl at the feed inlet end in order to take advantage of the colder feed temperature for cooling. To accpmpliish this, it is necessary to flood a portion of the tubes with steam oondensate, which results in poor ueilization of part of the heat transfer surface. When it is possible to do so, high-pressure feedwater heaters should be arranged horizontally aa illustrated by Fig. 33(a). Compared with a vertical position, a horizontal arrangement affords the following advantages: There is less restriction on length. A heater of longer length normally results in a smaller shell diameter and a more economical unit. The ieheat problem is less severe in the steam condensate-cooling section because this section need not be exposed to the steam and a short condensateoooling section can be used to achieve the proper proportion of the tube surface to shroud surface. w A higher condensing heat transfer coacient is achieved on a horizontal tube bundle than on a tube bundle that is vertical. Tbis fact is not reflected in most design analyses; however, it should be considered

(a) Torrgueand-Growe Flange Joint fig.

lbl Sheer-block d a r e

34 Tubedde cfoums

by the designer so aa to provide the most effective utilization of heat transfer surface. Of major importance in the design of high-pressum feed heaters is the adequacy of the c l o m s for the pressures and temperatures involved. A breakable joint should be provided for the shell-side c l ~ soethat the shell can be removed from the tube bundle for inspection and cleaning purposes. Under the usual operating conditions, a bolted flange joint ifil suitable for this purpose. However, under conditions which would involve temperature distortion or when there is an infrequent requirement for removal of the shell, a welded joint provides positive sealing lihd is economical to fabricate. A back-up ring is provided to protect the tubes when a flame cut is made to open the welded joint. The feedwater that enters the channel and the tube side of the unit is under a relatively high pressure which imposes a severe requirement on the tube-side closure and seal design. The bolta of a flanged joint are required to take the hydrostatic load (whieh depends on the closure diameter and fluid pressure) and at the same time maintain a pressme on the gesket s d c i e n t ts ensure a seal. This often results in huge bolts which require enormous torques to tighten, e q e d l y so when the larger shell diamet~?rsme involved with pressures over 1200 pig. Nevertheleae, bolted-flanged joints can be properly applied in the design of heaters with diameters as 1-e as 20 inches and for pnxmes of less than lux, psig. Flat metal or metal-jacketed gaskets are frequently used with bolted tube-ride closures. The force required to adequately compress the gasket is a mbt~dltiat percentage of the h y load ag~the end .clow~e ~ and may even exceed it. The force required toobtain an adequate gasket seat can be reduced by narrowing the width of the gasket and at the same time confining the gasket to an enclosed space to prevent the gasket from deflecting freely. An example of this design feature is illustrated by the tongue-and-groove hnged joint in Fig. %(a). A shear-block closure design that is used in some of the

MARINE ENGINEERING

larger and higher-pressure feed heaters is depicted by F i p . 34(b) and 35. The hydrostatic load is resisted by the shear ring, and the pressure on the gasket is maintained by the hydrostatic load. Figure 35 shows the method by which the tube-side operating pressure is used to seat and seal the solid copper ring gasket. With the larger high-pressure heaters, a point is reached where a simple bolted-flange closure must give way to a more elaborate high-pressure closure; experience indicates that when the product of the operating pressure (in psi) and the shell I D (in inches) exceeds 25,000, a bolted-flange joint is no longer economical. In actual design practice, the availability of a standard flange for the size and pressure intended often decides the question as to the specific closure design. The economic advantage of a ready-made versus a custom-made item will often influence the selection of the specific closure. 3.8 Gland Leak-Off Cqndensers. I n order to avoid an ingress of air into the steam system at the points where the steam turbine shaft penetrates the turbine casing, and similar locations, a gland-sealing steam system maintains a pure steam atmosphere at a pressure slightly above atmospheric just outside the turbine shaft-casing interface; this ensures that atmospheric air will not enter the turbine. The gland leak-off system consists of a fan which removes an air and steam mixture from the turbine gland leak-off pockets, and a condenser through which the mixture is drawn in order to condense the steam so as to recover the water and reduce the quantity of gas which the fan must handle. The gland leak-off condenser may be either furnished as a separate heat exchanger or combined as a section

HEAT EXCHANGERS

of the low-pressure feed heater (see Section 3.5). When it is a separate exchanger, it is of the U-tube design and is arranged to receive low-pressure steam in the shell, and cooling water is passed through the tubes to effect the necessary condensing of the steam. The fan which forms a part of the gland leak-off system is designed to handle the leakage air plus the uncondensed steam vapor. The gland leak-off exhauster (which is the common name for this fan) is usually mounted on top of or immediately above the condenser (or condenser section of the first-stage heater). The Navy has set forth specific design and material requirements for this condenser and similar condensers for other shipboard applications [17]. 3.9 Unflred Steam Generators. Unfired steam generators (which are also referred to as contaminated water evaporators or steam service evaporators) supply lowpressure steam at a pressure of 50 to 150 psig to a system that is independent of the main steam system. The independent system provides steam for services which could possibly contaminate the main system in the event of a system malfunction. Some of the "contaminated" services include the fuel-oil suction and service heaters, cargo-tank heating coils, galley and heating systems, and steam-driven deck machinery. The contaminated evaporator tube nest drains are normally piped to the deaerating feed heater via a trap. Figure 36 depicts a typical contaminated steam system. Contaminated evaporators operate on bleed steam from the high-pressure turbine, auxiliary steam, or in some cases, high-pressure steam (up to 500 psig). The bleed steam is a variable-pressure source which depends on the percentage of full power being developed. The requirements of the contaminated system also vary greatly. Operational requirements, as depicted on a normal-power plant heat balance, may be only a small percentage of the maximum performance requirements.

PRESSURE REDUCING (TYPICALLY 600 TO 150 PSIG)

SAFETY VALVE

FROM DESUPERHEATER'

, ' STEAM LINE

TO D.C HTR.

ORIFICE

*ONE OF MANY POSSIBLE FEED CONTROL ARRANGEMENTS FLOWS TYPICAL FOR NORMAL OPERATION 1

CONTAMINATED DRAIN TANK Fig. 36 Typical contaminated steam system

The maximum performance requirements must be taken as the design conditions with checks niade to ensure that the design is satisfactory for other operating arrange-

LOW PRESSURE STEAM OUTLET

SPLIT RETAINING RING

I N N E R COVER

POSITIONINGR I N G

(J SHEAR-BLOCK

TYPE CHANNEL COVER

Two sets of design conditions are generally set up: one for bleed-steam operation and one for auxiliarysteam operation. Both conditions must be considered to determine the effectson the evaporator design as well as the safety valves and orifices. The evaporator bundle is normally of the U-tube type as shown by Fig. 37. When the tube bundle becomes very large, straight tubes should be used, incorporating the outside-packed head type of construction illustrated by Fig. 38. Tubes are usually %-in. OD on a lxs-in. square pitch, except in the case of a low temperature difference when a closer pitch may be used (N in.). H-in.-OD tubes, when used, are placed on a 1%-in.

BLOWWW Fig. 37 Typical contaminated water evapaata

(b) CONSTRUCTION DETAILS OF SHEAR-BLOCK TYPE CLOSURE

Fiq. 35 shear-block closure

water gage glass, steam supply r e d a t i n g valve, ther-

The overall clean-tube heat transfer coefficients that

526

HEAT EXCHANGERS

MARINE ENGINEERING

I

are used to determine the evaporator bundle size are as

(T,*,, - t , ~ ) ... . . . . . . . Overall heat transfer coefficient .............

laO deg F

550

520

500

OVERFLOW

A steam flow orifice is usually installed in the steam supply line to limit the amount of heating steam entering the evaporator and reduce the steam supply pressure (a high heating steam pressure is not always beneficial due to the critical heat flux or the vapor blanketing phenomenon iri boiling). The orifice size is determined from the Fig. 38

=

105f[P.(Pl - P2)I1/'

if P2

Outside-packed head construction used on large tube bundler

> 0.58P1

a drain to the bilge. A probe for oil detection and an associated alarm are also normally provided. The filter section is packed with a loose filtering medium, usually loofa sponges, arranged for easy access and replacement of filter material. A, = orifice area, sq in. The storage section is provided with a gage glass, P1 = upstream pressure, psi thermometer, removable cover, cooling coil, a drain and P2 = downstream pressure, psi overflow to the bilge, and a vent. W#,,= steam flow rate, lb/hr 3.1 1 Hot-Water Heaters. Hot-water heaters are The safety valve should be sized for the maximum used to provide water at controlled temperatures for showers, lavatories, and galley services. Hot-water steam flow entering the generator, i.e. heaters are generally steam heated and can be classified into two types: storage heaters and instantaneous heaters. W = 105.3 Ao[P4(Pa - Pr)]1'2 A storage type of heater has the advantage that the water can be heated and stored during nonpeak periods of hot if P4 0.58Pa W = 51.45 AZa, water demand, thus reducing the peak heating-steam requirement. Also a less sophisticated temperature control device is required for a heating rate that is Pa = Pressure of line (Assume independent of the rate of hot water withdrawal. The if P2

< 0.58P1




BRINE BLOWDOWN

CONDENSATE

Flow diagram of a long-tube tlash evaporator, brine reci~ulationsystem

As indicated by Fig. 9, treated seawater is pumped through all of the condensing tubes, from the last stage to the first, enroute to the feedwater heater. As the feedwater flows through the condenser tubes, it is heated progressively in each stage by the vapors which condense on the tube outer surfaces. The feedwater then passes through the tubes of the feedwater heater where it is heated to its terminal temperature. The heated feedwater is discharged to the shell side of the evaporator first stage for flashing. The remaining brine flows successively from the first to the last stage, reflashing in each stage, and is pumped from the last stage back to the ocean. 1.4 Thin-Film Evapora?ors. Section 2 of Chapter 14 discusses details relative to the transfer of heat from one fluid to another through tube walls, and Fig. 11 of Chapter 14 illustrates the resistance to heat transfer encountered when transferring heat through a tube wall. By analyzing the temperature gradients involved with the transfer of heat through tube walls, it becomes apparent that if thin films can be created and maintained on the tube walls, a relatively higher overall heat-transfer coefficient can be maintained. Doubleflute tubes have been employed in several marine evaporator designs as a means of maintaining thin liquid films. The configuration of the so-called "thin-film" tube design employed is shown in Pig. 10. The arrangement of the double-flute tubes in the evaporator is such that the tube orientation is vertical, with steam condensing on the outside and a falling film of feedwater evaporating on the inside, as indicated by Fig. 11. The operating principle of the double-flute tube is such

EVAPORATOR

i

I

I

1

DISTILLATE

( 1

FEED HEATER

that the condensate which forms has surface tension forces acting to drain it from the crests into the grooves. This feature results in the major portions of each crest on the evaporating and condensing side having a very thin film of liquid to greatly reduce the resistance to the heat flow through the crest area. The liquids in the grooves are channeled off by gravity, and the heat flow through this area is somewhat less. Seawater is introduced on the inner surface of the tubes by a spray nozzle; and the falling film of brine on the evaporating side tends to collect in the grooves by surface tension. A secondary but important function is performed by the flutes in organizing and controlling the falling film to assure uniform distribution of the brine down the length of the tube. Two-effect thin-film evaporators, such as illustrated by Fig. 11, have been installed on a number of ahips. Such t h i n - h evaporators offer the advantages of good heat-transfer characteristics, easy tube removal (provided by O-ring joints), and low probability of distillate contamination due to brine carry-over (only a small quantity of brine is maintained in the evaporator). However, chemical feedwater treatment is required with these evaporators, and experience has shown that the feedwater chemistry must be diligently maintained in order to avoid a rapid accumulation of scale on the thinfilm heat-transfer surfaces. The "spray-film" evaporator is a popular form of the thin-film type of marine evaporators. The spray-film evaporator is especially adaptable for application in an "in-cycle" or "condensate-cooled'' distilling plant SYStern. Spray-film evaporators, employing condensate

t

1

t

DlSTl LLATE

BRINE

CONDENSATE

Fig. 9

Flow diagram of a long-tube tlash evaporator, once-through system

from the main propulsiori plant condenser as the coolant in the distilling plant condenser, have been installed aboard ship. With such an arrangement, the condensate-cooled distillirig plant accomplishes a secondary objective of functioning as a low-pressure feedwater heater in addition to its primary objective of producing potable and high-purity makeup feedwater from seawater. The combination results in a higher overall cycle efficiency [4]. A typical flow diagram of a condensatecooled spray-film evaporator is shown in Fig. 12. The heating bundle of a spray-film evaporator consists of a horizontal tube bundle with the heating medium on the inside of the tubes. The heating mediumcan be bleed steam from steam turbines or steam produced from waste heat from diesel engines or gas turbines. The brine level is maintained in a hotwell below the evaporator tube bundle and a brine pump provides the necessary pressure to recirculate the brine to the spray nozzle header. Recirculated brine from the spray nozzles is "rained" over the top of the heating surface where it flows from tube to tube in thin films, resulting in relatively high "thin-film" heat-transfer rates. An a u t e matic chemical feedwater treatment system is provided to retard the formation of scale on the heating surfaces. The vapor generated at the tube surface counterflows through the curtain of recirculated brine and fine liquid droplets are thereby removed from the vapor. Removal of the liquid droplets in this manner, coupled with the secondary entrainment separation in the demister column, enables the production of distillate having a very high purity. A steam-motivated air ejector or mechanical vacuum

HORIZONTAL CROSS SECTION

-

EVAPORATION SlDE

.

WALL

THIN FILM REGION

/

--

THIN FILM REGION

4

'CONDENSATE

HEAT FLOW CONDENSING SlDE

Fig. 10

Double-tlute thin-tllm heat-transfer surface

pump can be used to maintain a low shell pressure in the evaporator. The air ejector after-condenser can be a separate heat exchanger or it can be incorporated within the evaporator heating bundle, thereby utilizing the heat content of the air ejector steam without necessitating a separate after-condenser heat-exchanger assembly. Since a spray-film evzqm-ator opeates at relatively higher shell pressures arid lower feed rates than a marine once-through type flash evaporator, the vacuum equip ment handles a lower volume of entrained air and corrosive gases and at a lesser vacuum level. As a result, the application of vacuum pumps in lieu of air ejectors becomes more attractive. To ensure conformance with the regulations of the U. S. Public Health Service [5], the shell temperature

SEA WATER FEED

SECOND EFFECT HEADER \

n

54 1

DISI'ILLING PLANTS

MARINE ENGINEERING

540

/FIRST

DISTILLATE COOLER

EFFECT HEADER

ORING TUBE JOINT

BOILER FEEDWATER c3 INLET 135.000 LBIHR 910F 270 GPM 150 PSlG

LEGEND w

STRAINER I* SALINITY CELL

LP STEAM SUPPLY

SECOND EFFECT VAPOROUTLET

NONPOTABLE DIST. TO WASTE

SECOND EFFECT MESH SEPARATOR

FIRST EFFECT ESH SEPARATOR

GATE VALVE GLOBEVALVE NEEDLE VALVE CHECK RELIEFVALVE VALVE

DISTILLATE TO STORAGE WW LWHR 14 GPM

fp

GAUGEGLASS THERMOMETER

9

PRESSURE GAUGE TEMP. SWITCH

PSlA 4 5INLET PF BLEED 26.5 STEAM SECOND EFFECT VAPOR CHEST

81264 BTUILB 7330 LBIHR EXHAUST STEAM INLET DESUPERHEATER WATER 547 LBIHR 8 91°F 150 PSlG CONDENSATE TO 1ST STG. HTR.

FIRST EFFECT DOUBLE-FLUTE SECOND EFFECT DOUBLE-FLUTE TUBE NEST ORlNG TUBE JOINT DISTILLATE PUMP.

SECOND EFFECT BRINE SUMP

CH SECOND'EFFECT BRINE DRAIN

Fig. 1 1

EJECTORCONO. OUTLET 250 LB/H BRINE PUMP 160 GPM 35 P.S.I.

SECOND-EFFECT FIRST~FFECT FIRST~FFECT DISTILLATE DRAlN BRINE DRAIN STEAM CHEST DRAIN

Two-effect marine thin-tllm evaporator schematic SEAWATER INLET 41 GPM O( 850F 20850 LBIHR 40 PSlG PRESSURE

Fig. 12

must be maintained at a minimum temperature of 165 F so as to pasteurize the vapor and distillate. A temperature switch is provided to sense the shell temperature and actuate the distillate dump valve, diverting the distillate to waste should the shell temperature fall below the 165 F minimum. Where lower operating temperatures are desirable or necessary due to a low-temperature heating medium (such as engine jacket water), a separate distillate sterilizer must be incorporated. As indicated by Fig. 12, a blowdown cooler cools the hot blowdown while simultaneously preheating the incoming makeup feedwater. The brine concentration in the evaporator sump is maintained at the proper density by setting a continuous blowdown rate. A rotometer is provided to indicate the blowdown rate. A shell-and-tube type distillate cooler is used to reduce the temperature of the distilled water produced. Seawater is normally used as a coolant in the distillate cooler in order to reduce the temperature to a "potable" water temperature. Blowdown and distillate coolers have sometimes been combined into one plate-type heat exchanger utilizing a 3-fluid flow pattern, i.e., hot brine, hot distillate, and cold makeup feed. Condensate-cooled distilling units are often used in conjunction with a combined firstr and second-stage feedwater heater/gland exhaust condenser in a manner similar to that discussed in Section 3.5 of Chapter 14. The resulting packaged spray-film evaporator feedheater is illustrated in Fig. 13, which is a photograph of a unit ready for installation aboard ship. The advantages of compactness and increased operat-

ing economy of the packaged feedwater heater/condensate-cooled distilling plant are of significant importance from the viewpoints of machinery space arrangements and overall cycle efficiencies. 1.5 Basket Evaporators. Evaporators of the basket type are specifically designed to employ "cold shocking" as a means of preventing an accumulation of scale on the evaporator heating surfaces; also, the deeply corrugated basket provides a large amount of heatrtransfer surface in a given area as may be noted from the evaporator sectional view in Fig. 14. The diagram of a double-effect basket-type distilling plant illustrated by Fig. 14 shows that the cycle is commenced by supplying steam to the basket-type heating section of the first effect. The latent heat of the steam is transferred through the basket wall to the seawater and the condensate formed is returned to the boiler. A circulation baffle or "skirt" is located a fixed distance around the basket on the seawater side. As a result of the skirt, violent boiling action takes place between the skirt and the basket. The heavy foaming characteristic of the boiling seawater inside the skirt produces high velocities and a low static head, thereby reducing the pressure in the region of boiling. In addition, the high velocity over the basket surface has a scrubbing effect which tends to retard rapid scale formation. After leaving the evaporator section, the vapor passes through a cyclone separator and a "snail" where particles of brine entrained with the vapor are removed. From the snail the vapor passes through the feedwater heater and into the basket of the second effect. The

BRINE BLOWDOWN TO WASTE 27 GPM 120°F

-

.-.-

16 .- Pel(:

I

DISCHARGE PRESSURE 73800 LBMR

Condensate-cooled spray-tllm evaporator tlow diagram

j

latent heat is given up by the vapor and the condensate produced is discharged to the flash tank on the distiller condenser. The vapor produced from the seawater in the second effect passes through separators, similar to those in the first effect,and then goes to the distiller condenser where its latent heat is transferred to the seawater in the condenser tubes. After absorbing heat in the condenser, the majority of the seawater coolant is discharged overboard; however, a small quantity (equal to about three times the distillate produced) passes through the tubes near the top of the condenser where further heating takes quantity of seawater, now called feedwater, place. T h i ~ is discharged from the condenser to the feedwater heater where the feedwater is heated to within several degrees of the evaporating temperature in the first effect; finally the feedwater enters the bottom of the firstrstage shell. The excess feedwater (about twice the evaporation rate) maintains the required level of seawater in the shell. The entrained excess,feed is carried up with the vapor produced, but it js removed from the vapor in the separator and falls into a brine gutter in the bottom of the separator section. The brine collected is discharged to become feedwater for the second effect. The quantity of brine carried off from the first effect by entrainment is sufficient to feed the second effect and to maintain the correct brine density in the first effect.

t

I-

The distillate collected in the flash tanlc of the distiller condenser is pumped through the distillate cooler after which it is either discharged to a distillate storage tank or dumped to waste. Returning to the design of the basket itself, the basket is made in sections of 0.043 to 0.050 in. moilel sheet that are given a corrugated configuration in a press. The sections are then welded to form the basltet shape shown in Fig. 14. On the seawater side, the basltet surface is highly polished; this is done to minimize the tendency for scale to firmly adhere to it. In order to control the accumulation of scale on the basket surface, it is necessary to execute a cold-shocking procedure every 75 to 100 hours of continuous operation. To do this, the brine is drained from the unit and steam at about 15'psigis admitted to the basket heating section. The flat sides of the corrugations expand, and the 250 F steam dries the scale. Next, the steam line is secured and cold seawater is allowed to cascade down over the basket. As the steam insidsthe b a s k t condenses, the pressure inside quickly drops from 15 psig to-approximately 28 in. Hg vacuum, causing the flat sides of the basket to contract. The scale is consequently cracked off the basket and drops to the bottom of the shell where it is removed through the cleanout door. Repeating this process several times removes the majority of the scale from the heating surface. In addition, at infrequent intervals, cleaning with a solvent is recommended so as to

-

I

MARINE ENGINEERING

543

DISTILLING PLANTS SECOND EFFECT AIR OFFTAKE TO EJECTOR OR VACUUM WMP

1

HEATING SECT10

CIRCULATING WMP

LEGEND EQUALIZEA

STEAM

SKIRT

IST EFFECT VAPOR

CORRUGATED BASKETTYPE HEATING SECTION

ZND EFFECT VAPOR SEAWATER CONDEWSATE

Fig. 13

Spray-fllm evaporator

dissolve any accumulation of scale on the basket, heat exchangers, separators, and pipes. 1.6 Vapor-Compression Distilling Plants. Vaporcompression distilling plants are designed for service where low-pressure steam or diesel engine waste heat are not available in sufficient quantity to operate an evaporator. The major advantage of a vapor-compression evaporator is its high thermal efficiency and that it operates on a self-contained thermodynamic cycle that is dependent only upon a source of power to provide the input energy required. The power supply may be in the form of electrical energy to operate the electric boiler, motor drive for the vapor compressor, and motor-driven pumps; or the compressor can be driven by a diesel engne or gas turbine. When the compressor is driven by a diesel or gas turbine, the exhadst gases can be utilized as auxiliary boiler heat. Special arrangements can also be made to use steam heating coils for the boiler if a small amount of steam is available for this purpose. Most offshore drilling rigs are equipped with vaporcompression plants as are some gas turbine ships and submarines. The evaporator in a vapor-compression distillation plant can be any of a variety of designs. Possible types

of evaporators include the spray film, submerged tube, basket, or vertical type of tube bundle with seawater in the tubes. Figure 15, which is a diagram of a vaporcompression distilling plant that employs a spray-film evaporator, illustrates the principles involved with a vapor-compression distilling plant. Feedwater is pumped through a solenoid valve, control valve, flowmeter, heat exchanger, vent condenser tube bundle, and then into the spray pipe manifold from which it is sprayed over the tube bundle in the evaporator shell. Some of the sprayed brine striking the hot tube bundle evaporates into steam vapors which are drawn through the demisters into the vapor compressor. The brine that is not vaporized collects in the bottom of the evaporator shell and flows into the evaporator sump. The recirculation pump takes suction from the evaporator sump and returns the majority of the brine to the spray pipe manifold. Incoming feedwater is used as cooling water for the recirculation pump mechanical seal and is then combined with the recirculating brine. The recirculation pump also pumps a portion of the recirculating brine through the evaporator sump liquid level control to the blowdown side of the heat exchangers, and then to discharge. For seawater use, the blowdown

EVAPORATOR SHELL SECTIONAL VIEW OF EVAPORATOR

Fig. 14

Double-effect basket-type distilling plant

flow rate is normally set at twice the total distillate flow for best operation. The boiler section integral with the evaporator provides the small quantity of starting and makeup heat required for the operation of the distiller. If the boiler is steam heated, low-pressure steam is piped into a small U-tube bundle to provide the necessary heat. A lowpressure (15 psig) steam supply free of contaminants is required to maintain a compressor suction pressure of 0.5 psig and is usually regulated by a diaphragm-type control valve. If the boiler is electrically heated, three electric immersion heaters are generally used with two of them being manually controlled and the third one operated automatically. All three heaters are put in service during start-up to get the plant up to normal operating temperatures as quickly as possible. During normal operation, the manual heaters are used as required, along with the automatic heater, to maintain a compressor suction pressure of about 0.5 psig. The automatic heater is controlled by an evaporator shell pressure switch. A'fter start-up, some of the distillate is used as boiler makeup water. A float-operated control valve in the distillate circuit regulates the correct amount of makeup water into the boiler.

The vapors produced from the brine on the shell side of the evaporator are drawn through demisters into the vapor compressor. Vapor compressors of a centrifugal design, that operate at a relatively high speed and low noise level, are most suitable; however, many plants have been equipped with a single-stage, positive-displacement, three-lobe rotary-type compressor. The compressor normally operates with a 2 to 4 psi differential pressure between the suction and discharge; the maximum allorvable differential pressure is about 5 psi. During the compression process, the steam vapor increases in' pressure and temperature after which, in a spray film evaporator, it is discharged into the tube side of the evaporator tube bundle. The latent heat of the steam vapor is transferred through the walls of the tubes to the brine being sprayed, over the tube bundle.= This transfer of heat condenses'tffe steam vzpor inta distillate which flows out of the tubes into the bottom of the steam chest, A spray pipe assembly is the standard means whereby recirculating water is sprayed over the evaporator tube bundle. A steam chest and vent condenser are bolted to the evaporator tube bundle. The steam chest channels the flow of steam vapors from the compressor jnto the tube bundle and the flow of distillate from the tube

DISTILLATE S W A Y CONTROL VALVE ,

PRESSURE RELIEF I

bundle. Incoming feedwater flows through the tubes of the vent condenser, where it gains additional heat from the vapors condensing in the vent condenser section. The vent condenser is vented to the atmosphere to dis* charge noncondensible gases from the steam chest. The distillate pump takes suction from the steam chest and pumps the majority of the distillate through the heat exchanger, the flowmeter, and then to the discharge connection. A small portion of the distillate enters the boiler t h r o u b the boiler water level control valve as makeup water. Another small portion enters the compressor suction duct and serves as compressor desuperheating and sealing water. A spray-film vapor-compression distilling plant is furnished by the manufacturer as a package complete with all interconnecting piping, electric wiring, automatic controls, and insulation; such a unit is shown by Fie. 16. Over a period of time, scale-forming elements in the feedwater gradually accumulate on the evaporator tubes and lower the rate of heat transferred from the compressed steam to the recirculated water (scale control is discussed in Section 2.5). Normally the amount of steam compressed is constant. Therefore the compressor differential pressure rises (causing an increase in the temperature difference) to counteract the effects of the scale accumulation. 'As the compressor differential pressure rises, there will be a slight decrease in distillate production; therefore, an acid cleaning system must be provided and used when the rated distillate capacity can no longer be maintained or when the com~ressordifferential Dressure exceeds 4.5 psi. When the lpressure differentiil across the compressor reaches 5 psi, the unit is operating at its minimum rated capacity and the evaporator tubes must be cleaned to raise the distillate o u t ~ u t lower . the electrical energy input, and prevent overloading the compressor motor. When it is desirable to use a diesel engine instead of an electric motor to drive the com~ressoka conventional four-cycle, in-line, medium-speed, industrial-type engine is best suited for continuous operation. The engine directly drives the compressor and both water pumps via V-belts. Engine jacket water can be used as a heat source to the boiler thereby providing a means of returning the heat in the jacket water to the system. The boiler section can also use the engine exhaust gas as a heat source; by doing so, maximum use is made of the engine waste heat. Detailed material and design requirements for distilling plants of the vapor-compression type for Navy applications are contained in reference [6]. The plant installed on some submarines operates on the principle described in the foregoing, but the plant is completely electrically operated and incorporates a vertical-tube bundle as illustrated by Fig. 17. Referring to the plant shown by Fig. 17, the entering seawater is preheated (for arctic operation or cold feed due to deep submergence) in an assembly that consists of two cylindrical shells, each containing an electric

-

EVAPORATOR DESUPERHEATER

u

DISTILLATE OUT

FEEDWATER SOLENOID VALVE

1 2 Fig. 15

---

FEEDWATER IN

-.

BLOWDOWN OUT

Vapor-compression distilling plant with spray-fllm evaporator

Fig. 1 6

Spray-fllm vapor-compression distilling plant

545

DISTILLING PLANTS

MARINE ENGINEERING

immersion heater. The temperature control unit automatically energizes one or both of the heaters when the incoming seawater temperature is below 55 F and turns off the heaters when the temperature is above 65 F. A rectangular shell-and-double-tube heat exchanger, or one of similar features, is located in the line between the feedwater preheater and the yent condenser. The shelland-double-tube heat exchanger is vertically divided into two sections of equal size, a distillate side and a brine side. Inner anc)'outer tubesheets are bolted to each end cover plate. Larger, straight outer heat-transfer tubes are roller-expanded into the inner tubesheets and smaller straight tubes, inserted through the outer tubes, are roller-expanded into the outer tubesheets. Incoming feedwater flows in the annular space between the inner and outer tubes while the distillate and brine, discharge from the evaporator, flow through the inner tubes in their respective sides of the heat exchanger. The feedwater flow is counter to both the distillate and brine flows, providing good heat transfer between the flows. The unit serves to further heat the feedwater and simultaneously cool the distillate and brine. The feedwater leaving the heat exchanger next passes through the tubes of the vent condenser, where it gains additional heat as the hot noncondensible gases vented from the evaporator section are cooled. The hot feedwater leaving the vent condenser is discharged into the large circulation tube located in the center of the evaporating section and collects in the brine recirculation sectioq. .- The temperature of the brine in the recirculation section is raised to the boiling point by heat coming from the boiler section. The boiler section is located on the floor of the evaporator. Its source of heat is three electric immersion heaters. During operation the electric heaters keep the distillate at the boiling point, which in turn further heats the brine in the brine recirculation section located above the boiler section. Since the feedwater is preheated when it enters the evaporator, comparatively litt~e heat is required from the immersion heaters to keep the brine at the boiling point. A brine overflow tube, located in the center circulation tube, runs through the feed section to a sidewall connection. Two distillate drain tubes from the evaporator section to the boiler section also run through the feed section, astdoes a steam vent tube from the boiIer section to the evaporator section. When the brine boils in the recirculation section, it passes into the evaporator section tube bundle where more heat from the compressed vapor is applied. This action "o'f the ly-irie, and heating increases the boil* about two thirds of the brine is vaporized at a pressure of about 1 psig in the tubes. The brine not vaporized flows into the brine overflow and is piped out of the evaporator into the heat exchanger. Steam vapors from the boiling brine in the evaporator section tubes rise into the vapor section and are drawn into the suction side of the compressor through the baffle arrangement and demisters. The baffle arrangement and demisters remove *

MARINE ENGINEERING

DISTILI.ING PLANTS

entrainments from the vapor and allow relatively ('clean" vapor to flow to the compressor suction. The vapor compressor, which is the "heart" of the plant, is a centrifugal, liquid sealing ring type that is directly driven by a constant-speed a-c motor. The unit is mounted in the evaporator top cover plate and extends down into the vapor section. It operates at a suction pressure of about one psig at 215 F and a discharge pressure between 1.5 and 3 psig in the saturation temperature range of 217 to 222 F. When in operation, the compressor reqqres a continuous flow (about 0.25 gpm) of distilled water for sealing purposes. The seal water is necessary to provide proper suction and compression and to maintain the vapor discharge temperature within predetermined limits. The distillate discharge line, downstream from the heat exchanger, is tapped to provide the seal water. The compressor raises the temperature and pressure of the vapors and discharges the vapors into the evaporator section where a number of heat-transfer tubes are arranged in a pattern surrounding vertical baffle plates. The baffle plate arrangement directs the flow of vapors around the tubes and directs the noncondensable gases to a vent tube. A system of perforations in the vent tube collects the gases which are passed through an external vent condenser to the atmosphere. In the shell side of the evaporator section the vapors condense on the tube outer surfaces to form distillate which collects in the bottom of the evaporator section (on the lower tubesheet) and flows through the distillate return tubes into the boiler section. The excess distillate flows out through the outlet pipe into the distillate pump and is pumped through the heat exchanger and into the ship's storage tanks or to waste, depending on its purity. 1.7 Membrane Processes

a. Reverse Osmosis. Of the various membrane processes used in desalting, reverse osmosis has the most promise for shipboard applications. Developments in reverse osmosis have made membrane dpsalting processes widely accepted, and they are considered for seawater desalination. To understand reverse osmosis, it is necessary to review the basic phenomenon of osmosis. Osmosis depends on the existence of a membrane that is selective in the sense that certain components of a solution (ordinarily the solvent) can pass through the membrane, while one or more of the other components cannot do so. Such a selective device is called a "semipermeable membrane"; it is usually, though not always, in the physical form suggested by the word "membrane." As illustrated by Fig. 18, if a semipermeable membrane separates a solution from a pure solvent, or two solutions of different concentrations, the tendency to equalize concentrations will result in a flow of solvent from the less concentrated phase-that is, the phrtse richer in s o l v e n t t o t h e other; it is this flow of solvent that is termed "osmosis." I f an attempt is made to impede the flow by exerting pressure on the more saline solution (assuming for simplicity that the other phase is pure solvent), the

547

rate of flow will be decreased. As the pressure is increased, a point will be found at which the flow is brought to a complete stop, the tendency to flow being in equilibrium with the opposing pressure. This equilibrium pressure (actually, the equilibrium-pressure difference between the solvent and solution phases) is called the '(osmotic pressure"; the osyotic pressure is a property of the solution and cannot depend in any way on the membrane, so long as the latter has the necessary property of semipemeability. A further increase of the pressure on the solution causes reversal of the Ssmotic flow, and pure solvent passes from the solution, through the membrane, into the solvent phase; this phenomenon is the basis of the reverse-osmosis method of desalination. Most osmotic membranes have little mechanical strength and must be supported if they are to withstand large pressure differences. The development of supporting media to provide the necessary strength without seriously impeding flow is essential to the practical application of reverse osmosis. The most common semipermeable membrane used in reverse osmosis plants is cellulose acetate, the acetate ester derivative of cellulose. This is a chemical modification of cellulose in which some of the hydroxyl groups are replaced by acetate groups. Water selectively dissolves into the membrane and is transpoi-ted through the membrane by pressure-motivated diffusion. This is called "solution-diffusion." The membrane will reject trivalent ions better than divalent ions and divalent ions better than monovalent ions,. ,,Dissolved gases tend to pass through the membrane with very low rejections. Membranes have been produced and tested in such various forms and configurations as flat membranes (called plate and frame membranes), spiral or rolled membranes, and the tubular concept. The tubular membrane concept has emerged as one of the most popular due to its moderate space utilization, ideal pressure containment, light weight, ease of cleaning and servicing, and ease of membrane replacement. 'J'he application of reverse osmosis to shipboard desalting of seawater poses a somewhat different set of circumstances than those encountered on fixed-base installations. The variation of feedwater temperature from 28 to 85 F presents system design complications, because the osmotic membrane is sensitive to the water viscosity variation that occurs with temperature change. Permeation of solvent through a cellulose acetate membrane varies approximately 1.56 percent per degree F from a standard design temperature of 77 F. This change in rate of product water is partially offset by the change in average brine concentratio* As the membrane permeation rate is increased, with a constant feed rate, the salinity of the brine being circulated is also increased. The average brine concentration is defined as the initial brine concentration plus the effluent or waste brine concentration divided by two. For every 1000-ppm salinity change of the average brine in the unit, permeation through the membrane will vary 1.8 percent due to the change in osmotic pressure of the brine.

MARINE ENGINEERING

DISTILLING PLANTS

""U When fluids of different concentrations in a venoel are separated by e membrane, the dilute solution will flow thmugh the membrane into the concentrated solution.

"""U

1ST STAGE

Q MEMBRANE BANKS

9

2ND STAGE

) MEMBRANEBANK

FRESH WATER

F!!l

OSMOTIC PRESSURE The level of the dilute solution dmpr end the level of the concentrated solution rims until an "equilibrium" is reached The pressure difference between these two levels is the "osmotic pressure."

/

BRACKISH WATER

POTABLE WATER

WATER

I

MEMBRANE

SEA WATER

Fig. 2 0

REVERSE OSMOSIS

of the ounatic I f a pressure in expressure is applied to the concentrated solution, the flow is reversed from the concentrated solution to the diluted solution. This is "reverse osmosis." FRESH WATER Fig. 18

Fig. 1 9

I MEMBRANE

Osmosis, osmotic pressure, and revene osmosis

80,000-gpd barge-mounted reverse-osmosis desalination plant

Seawater desalination by revene osmosis

The first large reverse-osmosis plant to go to sea was an 80,000-gpd plant supplied to the U. S. Navy in 1969; it is depicted in Fig. 19. The plant is of the two-stage design employing a tubular cellulose acetate membrane. The membrane is "cast" inside a porous composition paper tube which is housed in a linearly grooved plastic support tube. The support tube is contained in the "tube-sheets" by means of an "O-ring" seal not unlike a shell-and-tube heat exchanger. Seawater desalination by reverse osmosis to produce potable water that meets the requirements of the U. S. Public Health Service [5] has been dome in two stages as indicated by Fig. 20. The initial feedwater containing. approximately 35,000 ppm of total dissolved solids (tds) is reduced to about 3500-ppm tds in the firstrstage permeate and thence from 3500-ppm to 350-ppm tds in the second-stage permeate. Research to develop a "single-pass" membrane and laboratory tests on cellulose diacetate, cellulose triacetate and cellulose acetate butyrate show promise that the "ideal" membrane to produce potable water (under 500-ppm tds) from seawater in one pass will be developed. Reverse osmosis will then challenge the evaporative processes for the lead in seawater desalination for potable use. Additional "polishing," such as the addition of a polishing demineralizer in the permeate stream, would be necessary to produce boiler makeup quality water (under 1-ppm tds). b. Electrodialysis. Whereas other forms of desalination schemes entail the removal of the vastly greater bulk of water from the salts, in the electrodialysis process the salts are removed from the water. As illustrctted by Fig. 21, with the electrodialysis (ED) process, an electrical field is imposed on the water by positive (anode)

SALINE WATER IN

WNC

IATED BRINE WASTE

I

-

1

FRESH PRODUCT WATER

Fig. 21

Schematic diagram of the electrodialysis process

and negative (cathode) electrodes. This field forces the positive ions to move toward the cathode and the negative ions toward the anode. An ED desalting "cell" is formed by two smooth, rigid plastic membranes with a spacer between, which guides the water flow. One membrane carries a permanent electrical charge that is positive and the alternate membrane has a negative charge. Just as magnets of like charge repel each other, the positively charged membrane W*els positi3 ions, yet ljermits negative ions to pass through it. The membranes and electrodes are so arranged to permit ions to leave an ED cell but none to return. Therefore, water in one c a m partment of an ED cell is desalted while the adjacent "brine" compartment is made more salty. Electrodialysis plants have successfully refined brackish waters of up to 2500-ppm total dissolved solids to produce potable water in shoreside installations [7] but have not been

I

MARINE ENGINEERING

550

proven to be economically feasible for converting highly brackish or seawater (feedwaters in excess of 5000-ppm tds) to potable levels. The cost of operation of ED plants is in direct proportion to the amount of salt removed. Since the salts actually pass through the membrane, the problems of fouling or "polarization" can be much more severe than encountered in reverse-osmosis systems. The use of cation-neutral membranes in ED systems lessens the danger of fouling. The operation of electrodialysis plants at temperatures

DISTILLING PLANTS

up to 180 F has been considered. At such temperatures, the electrical resistance to the flow of current through a solution is considerably decreased, resulting in lower power requirements. A lower power consumption could place ED systems in an acceptable cost range for processing the more brackish waters, and seawater, for marine applications. Other membrane processes such as ultrafltration and piezodialysis show little promise of being able to economicauy convert the more saline waters to a potable level.

Section 2 Distilling Plant Design Considerations 2.1 Heat Transfer in Distilling Plants. The fundamental theory of heat transfer in heat exchangers and the application of this theory are discussed in Chapters 2 apd 14. The basic analytical relationship employed in the design of heat-transfer equipment is

Q

=

UA(LMTD)

where quantity of heat transferred, Btu/hr heatrtrmsfer coefficient, Btu/hr= sq ftrdeg F = heatrtransfer area, sq f t A LMTD = logarithmetic mean temperature deg F =

Q

U

The main consideration in the thermal design of an evaporator is that fluids in the evaporator undergo a change in phase. In a flash evaporator, beat is transferred to the seawater in tubes by condensing a vapor on the outside of the tubes. In a submerged-tube or sprayfilm evaporator, the heat i s trwsfemed from the condensing steam inside the tubes to the boiling seawater on the outside of the tubes. AB in the design of heat exchangers (see Chapter 14), the overall heat transfer coefficient, U, is given by 1 U= (2) (I/&) 9.a rw rdt~ (l/h@) where

+ + + +

h,

=;

reciprocal

shelhide film coefficient

resistance of deposit or scale on tube outside wall r, = resistance of tube wall metal rdto = resistance of deposit or scale on tube inside wall - = reciprocal of tube-side flp coefficient

r*

=

values of heatrtransfer coefficients encountered in service are presented in references [8] and [9]. The overall heattransfer rate of a submerged-tube or thin-film evaporat~rdepends upon the vapor pressure of the liquid being evaporated, the temperature difference between the condensing saturated steam and the vapor (which fixes the steam pressure for a given vapor pressure), the disposition of the heating tubes in the shell, and the character of the liquid being evaporated as well as the cleanliness of the heatrtransfer surfaces. Heatr transferrates increase with increasing vapor and steam temperatures, with the temperature difference remaining unchanged, except that for seawater evaporation at temperatures exceeding about 200 F and for some classes of raw fresh water at the higher temperatures (particularly for high-temperature differences), scale forms so rapidly as to offset the advantage that would otherwise be gained. Under laboratory conditions with a single acid-cleaned tube re-evaporating distilled water, the temperature difference has a pronounced effect on the heattransfer rate. Rates varying from about 1300 Btu/hr-sq ftdeg F a t a 20 deg F temperature difference up to over 3000 Btu/hr-sq ftrdeg F at a 100 deg F temperature difference have been observed with the vapor at atmospheric pressure; the heatrtransfer rate falls off with further increases of the temperature difference. However, for scale-producing feeds, the scale resistance is such a large part of the total resistance to heat flow that the effect of temperature difference on continuous service rates is small. .Since t,he scale resistance is generally such a large ~ercentakeof the total resistance to heat flow for disiilling it is convenient to group all of the other together and express equation (2) in the folbwing form: 1 1 -+r Ue

us-

ho

All of the resistances listed in the foregoing are consistently based on the same area; by convention, the tube outside area L usually taken as the base. Typical

where

U,

clean-tube overall heatrtransfer coefficient; i.e., sum of film resistance on tube inside and outside surfaces and resistance of tube wall, Btu/hr-sq ftrdeg F r = fouling resistance; i.e., resistance of tube depo&itsor scale, hr-sq ftrdeg F/Btu Figure 22 is a plot of the clean-tube overall heatr transfer coefficient, U,, versus tube velocity that can be used in the design of a saltwater heater or condenser for a flash evaporator. The correction factor which must be applied to the coefficient read from Fig. 22 to compensate for inlet water temperatures of other than 70 F is given in Fig. 23; and correction factors for the tube material and gage are given in Table 1. Typical values of the fouling factor are given in Table 2. For a stage condenser in a flash evaporator, the fouling factor, r, will typically vary from 0.0005 to 0.001 hr-sq ftrdeg F/Btu. In the saltr water heater where elevated temperatures are encountered, the fouling factor ail1 normally vary from 0.001 to 0.0015 hr-sq ftrdeg F/Btu; see Table 2. A comparison of overall heat-transfer coefficients and temperature differencestypical of flash, submerged-tube, and spray-film distilling plants is shown in Table 3. Heatrtransfer data for a submerged-tube evaporator are usually expressed as a relationship between the heat flux, &/A = UAT, and the temperature difference, AT, =

55 1

because in a boiling evaporator the heatrtransfer coefficient is also a function of the driving force or temperature difference. Most available data are based on experience and therefore include allowances for fouling. Figure 24 is a curve showing typical values for submerged-tuhe evaporators. I n the case of a condenser such as encountered in a flash evaporator, the ternperdure difference between the fluids varies from location to location within the condenser in a manner similar to the temperature distribution in a heat qx'changer. There is a difference, however, in that the heat transfer for the condensing fluid takes place at a constant temperature and only the temperature of the liquid varies. For a condenser, the logarithmic mean temperature difference (illustrated by Fig. 25) can be expressed as:

where

Table 1 Heat-Transfer Rate Tube Material and Gage Correction Factor TUBE MATERIAL^ Admiralty metal Arsenical copper Aluminum Aluminum b r w Muntz metal Aluminum bronze 90-10 copper-nickel 70-30 copper-nickel Type 304 stainless steel

-TUBE WALLGAQE17

16

15

Bwg

Bwg

Bwg

0.97 0.97 0.97 0.93 0.93 0.87 0.87 0.80 0.56

0.93 0.93 0.93 0.89 0.89 0.84 0.84 0.76 0.54

0.88 0.88 0.88 0.84 0.84 0.79 0.79 0.71 0.51

2

3

4

5

6

VELOCITY, FEET PER SECOND

Fig. 22

Condenser clean-tube heat-transfer rate

Rg. 23 Hwt-transfer rate inlet temperature correction factor

(3)

TEMPERATURE OF INLET WATER. DEGREES F

7

8

MARINE ENGINEERING Table 2

DISTILLING PLANTS

Seawater Brackish water Cooling tower and artificial spray pond: treated makeup untreated City or well water (such aa Great Lakes) Great Lakes River water : mnlmum Mississippi Delawa~e,Schuylka East Rlver and New York Bay Chicago Sanitary Canal Muddy or silty Hard (over 15 grainslgal) ~ n g i n ejacket Distilled Treated boiler feedwater Bbiler blowdown

1

1

0.003 0.003 0.001 0.0005 0.001 0.002

0.002 0.003 0.001 0.0005 0.0005 0.002

0.004 0.005 0.001 0.0005 0.001 0.002

0.003 0.005 0.001 0.0005 0.001 0.002

the heating medium of 240-400 F. If the heating * Resistances are based on a temperature ofcooling medium is known to scale, these res~stances

IS

medium temperature is over 400 F and the should be modified accordingly.

Table 3 Heat-Transfer Coefficients and Temperature Differences Typical of Feedwater Heaters in Various Types of Distilling Plants

450300 300-650 500-800

5-50 15-45 10-30

T,

= vapor temperature 1, = cold liquid temperature k = hot liquid temperature

In a flash evaporator there are a number of temperature losses which must be taken into account in the design of the equipment. These losses are associated with the brine equilibrium temperature deviation, the boiling point rise, the demister pressure loss, and the condenser pressure loss. The effect of these losses on the LMTD is shown in Fig. 25. The losses are defined as follows: Brine Eauilibrium Temperature Deviation (DEV). The amount of superheat remaining in the brine as a result of incomplete flashing. Boiling Point Rise (BPR). The elevation of the saturation temperature of the liquid above that of pure water caused by the concentration of salts in the brine. Demister Pressure Loss (DEM). The equivalent ~A

20

25

30

40

35

46

55

50

TEMPERATURE DIFFERENCE, OF

I

Flash Submer ed tube spray fifm

553

Fouling Resistances Typical of Various Types of Water, r, hr-sq ft-deg F/Btv

saturation temperature loss associated with the pressure loss through the demister. Condenser Pressure Loss (COND). The weighted saturation temperature loss associated with the pressure loss through the condenser. I n marine applications, the losses associated with the brine equilibrium temperature deviation and the condenser pressure loss are usually small and c m be neglected; however, they are usually significant factors in higher capacity and economy plants. Figure 26 is a curve of boiling point rise versus temperature. Since the demister pressure loss is normally less than 0.5 deg F, it is common to allow 1.5 deg F for the combined demister l n n ~and *"---- - boiling point rise loss in the design of oncethrough marine-flash evaporato~. Because the operating brine concentration is higher in submerged-tube and spray-film evaporators, a value of 2 deg F is usually allowed for the combined loss for such units. I n order to illustrate the principles involved in the determination of the size of a condenser and saltwater heater for a distilling plant, consider the two-stage, 8000-gpd, flash evaporator illustrated by Fig. 5. The condenser temperature differences are computed as follows: STAGE1 STAGE 2

F.

Brine temperature, deg . : . . . . . . . . . . . D e m t e r and boiling ~ nrlse t loss, deg F vapor temperature, F . . . . . . . . . .. . . Condenser inlet temperature, deg F . . . . . Condenser outlet temperature, deg F . . . . Inlet temperature difference, deg F . . . . . Outlet temperaturqdifference,deg F. . . .

C

Fig. 24

Heat flux vs. temperature difference for submerged-tube evaporators

Since the temperature differences are the same in both stages, the two tube bundles will be made to the same design. The LMTD in the two condensers will be

LMTD = 33 - 8.25 = 17.85"F 33 In 8.25 The condenser tubes selected are %-in. o.D., 18-BWG (0.049-in. wall), 90-10 Cu-Ni tubes, and the tube design velocity is 5.5 fps. From Fig. 22 the heat-transfer rate for a % in. tube with a 5.5-fps tube velocity is 633 Btu/hr-sq ft-deg F. Applying an inlet temperature correction factor of 1.06 (since the condenser for the two stages will be the same, the correction factor read from Fig. 23 is based on the lower inlet temperature of 86 F) and a material correction factor of 0.90 (read from Table 1) gives a corrected clean-tube, overall heattransfer coefficient of 604 Btu/hr-sq ft-deg F. With a fouling resistance of 0.000675, the overall heat-transfer coefficient is computed from equation (3) as

U

=

1 1 604

+ 0.000675

=

FEED, T i ----c

TI,

+ + ----+ -------+ ---- ----DEV.

Tb' #

Tv Tv

f

-----,--,CONbAp

429 Btu/hr-sq ft-deg F

With a tube I.?. of 0.527 in., the number of tubes required to pass 60,000 lb/hr of feedwater having a density of 64.1 lb/cu ft at a velocity of 5.5 fps is

fl

Fig. 25

Temperature losses in distilling plants

DISTILLING PLANTS

MARINE ENGINEERING

Consequently, the required tube length is tube length =

0.5

60

100

150

200 TEMPERATURE OF

260

350

300

NOTE: THE NORMAL SEA WATER CONCENTRATION USED IN THIS CHART HAS 34.483 G SOLIDS PER 1000 G SEA WATER.

Fig. 26

Boiling p i n t rke vs. temperature

The tubes selected are %-in., 18-BWG, 90-10 Cu-Ni tubes and the design tube velocity is 6 fps. Reading a clean-tube heat-transfer coefficient of 660 Btu/hr-sq ftr deg F from Fig. 22 and applying an inlet temperature correction factor of 1.1 from Fig. 23, and s tube material correction factor of 0.9 from Table 1 gives a corrected For a feedwater flow of 60,000 lb/hr and specific heat clean-tube heat-transfer coefficient of 653 Btu/hr-sq ftr of 0.96 Btu/lb-deg F, the temperature difference of deg F. With a fouling resistance of 0.0015, the gverall 24.75 deg F corresponds to a heat transfer of 1,426,000 Btu/hr in each of the two condensers. This being the heatrtransfer coefficient becomes case, the tube area required is 1 = 330 Btu/hr-sq ftrdeg F U = 1 0.0015 653 No. of tubes required =

+

The number of 0.652 in. I.D. tubes required to pass 60,000 lb/hr of feedwater having a density of 64.1 lb/cu ft without exceeding a 6-fps velocity is

and the required tube length is tube length =

A

w

nm.

No. tubes required =

Using a 6-pass (3 U-bend) design, the required bundle length is determined to be 36.8/6 = 6.13 ft. The &e of the saltwater heater is established in a similar manner. The temperature of the condensing heating steam is 198 F and the saltwater inlet and outlet temperatures are 138 F and 170 F respectively; therefore the LMTD is

LMTD

'60 - 28 6 0

= -= In

28

42'F

(3600) (64.1) (6)

5 (0.6W2

No. tubes used = 19 The feedwater is heated from 135.5 F to 138 F in the air ejector after condenser. Since the feedwater has a specific heat of 0.96 Btu/lb, the heat transferred to-the feedwater in the saltwater heater is Q = (60,000) (0.96) (170-138)

=

and the heating surface required is

1,843,000 Btu/hr

A

-

?rDn

With a &pass design, the bundle length becomes 35.6/6 = 5.93 ft. 2.2 Heat Sources. The thermodynamic design of a distillation unit is strongly dependent upon the heat energy source to elevate the temperature of the incoming seawater feed to that required for efficient vaporization. This energy can be supplied as steam, electricity, or waste heat in the form of exhaust gases or hot water. A distillation unit cab use either, or any combination, of these energy sources. The selection of the optimum heat source is dependent upon the total plant heat balance, desired evaporator efficiency,availabIe energy, and mode of operation contemplated. Many ships are powered by steam turbines and, therefore, steam is a heat source commonly used in shipboard distillation units. Low-pressure turbine extraction steam and auxiliary exhaust steam are the normal sources of heating steam for distilling plant operation. To permit the most efficient and flexible operation, it is standard practice to design the distiller to operate with either steam that has been bled from the main turbine or steam that has been exhausted from any of the various auxiliary turbines. When operating at sea, bleed steam permits the most economical operation as the majority of the heat remaining in this steam is otherwise rejected to the main condenser, contributing very little additional energy to the cycle before being condensed (see Section 3.4 of Chapter 18 for further discussion regarding bleed steam). Ror in-port operation auxiliary exhaust steam is supplied to the distiller. The use of this reliitively low-cost steam permits continued economical operation of the distilling plant. The pressure of bleed steam is dependent on the turbine design and the specific extraction point. Distilling plant designs are normally predicated on steam being supplied at a pressure of 9 to 11 psia. Auxiliary exhaust steam is usually furnished at a pressure of 15 psig and reduced to 5 psig by a pressure-regulating valve; a pressure of 5 psig is usually provided at the inlet to the critical-pressure orifice. High-pressure air ejector steam, where employed, is also used to preheat the seawater feed. Heat in the air ejector steam is reclaimed in the air ejector aftercondenser using distilling plant feedwater as the coolant. The air ejector motive steam pressure is dependent on the air ejector design and can range anywhere from 75 to 150 psig; pressures in the 135- to 150-psig range are most common. Diesel or gas turbine driven ships often employ waste

555

heat as a heat source for~distillingplant operation. The waste heat may be in the form of diesel engine jacket cooling water or it may be in the form of heat recovered from the turbine or engihe exhaust gases. Submergedtube and thin-film distilling units have been used when hot water is the heat source because these units operate with a lower shell temperajure and can use water at a lower temperature as a heat source than can distilling plants of the flash type. Distilling plants for this application are~commonlydesigned to use jacket water with an inlet temperature to the distiller of 165 to 170 F. With a flash evaporator, the hot-water inlet temperature to the heater must be at least 190 F to obtain an adequate temperature differential to pennit an efficient saltwater heater design. A supply of diesel jacket water above 180 F is not normally available; however, water at such temperatures is obtainable on ships driven by gas turbines by installing heat exchangers to recover heat from the exhaust gases. See Chapter 2 for further discussion on this subject. Electric power is often used to operate the efficient vapor-compression type of distilling plants; however, due to considerations of economy, electric power is not normally used with distilling plants of other designs except as a supplemental or emergency heat supply. For flash, submerged-tube, or basket designs, the direct use of electrical power for feedwater heating is not practicable. For some applications, it is advantageous to use a combination of heat sources to operate the distillation unit. Figure 27 illustrates a typical flow diagram for a unit using all three common heat sources; this design has been applied in a number of diesel-powered naval ships. For the unit shown in Fig. 27, the primary heat source is diesel propulsion engine waste heat in the form of engine jacket water. This primary heat source is used when the ship is at sea operating with sufficient engjne power to provide the necessary waste heat. When the ship is operating at reduced power, the engine waste heat is augmented by heat furnished from either the steam heater or electric heater. When operating in this condition, the jacket water is directed through these heaters, thereby increasing the feedwater temperature. The steam heater is a shell-and-tube heat exchanger and the electric heater uses immersion heating elements: When the ship is in pee, all heat required to make rated capacity is supplied by the steam heater and electric heater. The steam heater is often sized to provide sufficient heat to produce rated capacity and the electric heater is sizedefi produce"'% rated cipacity. Aside from the increased reliability and operational flexibility provided by this combined heat source arrangement, an additional advantage is that the steam heater and electric heater are used to warm the diesel engine prior to starbup. 2.3 Distilling Plant Economy. The generally accepted meaning of the term "distilling plant efficiency" is the pounds of distilled water produced per 1000 Btu of

MARINE ENGINEERING

tions for ships constructed under the puwiew of the U. S. Maritime Administration are similar [ l l ] and 90-10 copper-nickel is also required for the flash chambers, with the provision that steel evaporator chambers, suitably protected against corrosion by means of a protective coating, will be considered. Protective coating materials provide definite savings in initial cost but the protection they provide steel surfaces in a marine distilling plant is sometimes iinsatisfactory. It is significant that pradt,ide in the use of nonferrous materials for distilling plants in ships built abroad followed that of the United States. Prior to 1950, distilling plants built in England and Germany used cast iron as the material for the shells. If the shells were thick enough, they were generally satisfactory, but the units were much heavier than when more corrosion-resistant materials such as copper-hickel, monel, and titanium were used. Not only is it important to use nonferrous inaterial in the areas in contact with the brine, it is equally important that the condenser baffles and distillate troughs and vent sections be of nonferrous materials. These come in contact with the vapolg which are laden with oxygen and carbon dioxide released from the heated brine and also are subject to corrosion. Due to rising costs and long lead times experienced in obtaining copper-nickel tubing, the use of titanium tubes in marine evaporators may become more common. See Section 5.6 of Chapter 22 for additional details relative to the materials of construction for marine distilling'plants and heat exchangers.

R P

CONDENSATEOUT D dirtlllats V vapor Jw FDW feed-

w-

Fig. 27

Flow diagram of an evaporator having three heat sources

heat supplied. For multistage plants, fitted with such features as evaporators which are heated by vapor produced in an earlier stage, the overall economy of the plant may be estimated best by first preparing an approximate heat balance flow sheet. Such flow sheets are approximate, but the balance a plant will assume in sewice can be predicted with fairly good accuracy. There are several factors which complicate a precise prediction of the manner in which a plant will perform in sewice. A "clean" plant is capable of producing excessive quantities of vapor to such an extent that carry-over will result, thereby contaminating the distillate; jn such an event the steam or feed supply must be throttled or other means used to control the output. Tube surfaces of different units foul to a different degree; therefore, fouling cannot be predicted with exactness. Additionally, the circulating water temperature varies with the season and locality. All of these factors affect the balance and economy of a distilling plant. The overall economy of a distilling plant, as distinguished from the so-called efficiency, may be expressed in terms of the pounds of distillate produced per pound of additional fuel required for the distilling plant, over and above the fuel otherwise required to fulfill the ship's power requirements. Establishment of the distilling plant overall economy entails an assessment of the amount of fuel used to produce the steam required for the saltwater heater, the electrical power required for the various pumps in the, plant, the useful work that could have been performed by the heating steam had it been used for other purposes, and the heat returned to the boiler feed system from the distilling plant (a very important consideration for distilling plants which have

557

DISTILLING PLANTS

condensers that use the main propulsion plant condensate as the coolant). Since the overall economy of a distilling plant depends upon the ship's power plant, and the interrelationships between the two, it is not possible to accurately state the economy of a distilling plant independently of the power plant. However, the following gives an approximate idea of the overall economy characteristic of four types of distilling plants: Single-stage submerged-tube plants operating on boiler steam at reduced pressure, 13 pounds of distillate per pound of additional fuel. Two-stage flash plants operatipg on bleed steam, 50 pounds of distillate per pound of additional fuel. Three-stage flash plants operating on bleed steam, 75 pounds of distillate per pound of additional fuel. Spray-film vapor-compression plants with electric motor drive and electric heaters, 200 pounds of distillate per pound of additional fuel. Based op flash distilling plants operated with exhaust or bleed steam, the weight of additional fuel required, together with the operating weight of the distilling plant, is bpt a small fraction of the weight of the distilled water produced during an average ship voyage. For a given ship, investigations may show t h a t a worthwhile savings in tonnage can be effected by using distilled water rather than tank water; and that, all things considered, the cost of distilled water is lower. 2.4 Materials of Construction. The U. S. Navy has standardized on 90-10 copper-nickel as the material used for tubes, tubesheets, waterboxes, evaporator shells, and piping for marine distilling plants [lo]. The specifica-

2.5

Scale Control and Acid Cleaning

a. Scale Control. High-temperature, land-based evaporators-have a much inore serious scale control problem than do marine plants which operate at a low temperature and normally under vacuum. Nevertheless, the formation of scale is a major consideration in the design of marine distilling plants. One source of scale is the bicarbonate in seawater which decomposes when the seawater is heated and then reacts with the magnesium and calcium in the seawater to produce magnesium hydroxide and calcium carbonate scales as shown in the following reactions: 3

1

2 bicarbonate ions

=

water

+ carbon dioxide gas + carbonate ion

i

carbonate 'ion

+ water = carbon dioxide gas + 2 hydroxide ions

+ calcium ion = calcium carbonate scale 20H- + Mg++ + Mg(0H)Z 2 hydroxide ions + magnesium ion = carbonate ion

magnesium hydroxide scale

When an acid or acidic salt is added to a seawater supply, it will neutralize what is commonly termed the alkalinity of the seawater. The alkalinity is the sum of the hydroxide, carbonate, and bicarbonate ions in the seawater, although usually only bicarbonate ions are present in significant ,quantity. The hydrogen ions produced in the seawater by tfie added acids or acidic salts act to destroy or neutralize the bicarbonates by the following reaction : ~

~

bicarbonate ions

0

+ H+ 3

-

+

COz

+ hydrogen ion

t

+ Hz0

=

carbon dioxide gas

+ water

Therefore, by treating the seawater with acid and converting the bicarbonates to form carbon dioxide and water, it is possible to eliminate the source of carbonate ions and prevent the formation of magnesium hydroxide and calcium carbonate scales. Such treatment is common practice in the operation of land-based multistage desalters. During the acidification of normal seawater, for scale control purposes, approximately 100 ppm of carbon dioxide is formed through bicarbonate alkalinity breakdown. This."carbonated brine" is acidic and corrosive to fermus materials of construction. The carbon dioxide can be removed by scrubbing the seawater with air or steam. Calcium sulfate scales are ~grobablythe worst of the common scales which may bk encountered when seawater is heated and evaporated to form concentrated brine. Calcium sulfate scales are not readilv soluble in acid solutions; consequently, they must often be removed by mechanical methods. The mechanism of calcium sulfate scale forination is simply precipitation caused by the concentration of calcium and sulfate ions beyond the solubility of calcium sulfate. Two crystal forms of calcium sulfate are involved, the anhydrite and hemihydrate. The anhydrite is the most insoluble form; therefore, its solubility is first exceeded uDon concentration of a feedwater containing calcium and sulfate ions. Solutions supersaturated with respect to the anhydrite are stable for long periods of time. Supersaturation with respect to the hemihydrate is not as stable as that for the anhydrite, so that scaling does not usually take place until the concentration of calcium sulfate hemihydrate is exceeded. The solubility of both the anhydrite and hemihydrate decreases with increasing temperature. The most common w & w f preverrting calcium sulfate scaling in marine seawater evaporators is toMm&ntaina blowdown rate sufficiently high that the brine concentration will not be saturated or supersaturated with respect to calcium sulfate hemihydrate. With proper distilling plant operation, calcium solubility limits are avoided and no calcium sulfate scales will be formed. Both thin-film and submerged-tube low-temperature marine distilling plants incorporate a chemical feed treatr ment system. This feed treatment system basically

MARINE ENGINEERING

DISTILLING PLANTS

I

consists of a chemical mixing tank and a proportioning pump. A small amount of plyphosphate is automatically added to the feed stream so as to chemically react with scale-forming ions to produce a soft sludgetype scale that is more readily washed from the tube surfaces than the calcium carbonate and magnesium hydroxide scales which would otherwise be formed. b. Acid Cleaning, During the operation of a marine seawater distilling plant, scale will form on the heating supfaces of the evaporator as reviewed in the foregoing. This scale is usually 80-90 percent calcium carbonate. The remainder is a mixture of calcium sulfate, magnesium hydroxide, metal oxides, silica, and miscellaneous deposits. However, when seawater feed is improperly treated with polyphosphates, other deposits such as phosphate sludges may be formed in large quantities. If fresh feedwater is used, silicates or calcium sulfate may be present. The rate of buildup and the composition of the deposits depend on such factors as the operating temperatures, the brine density, and flow rates. Feed treatment and cold shocking reduce the buildup of scale, but do not prevent or remove deposits entirely. The symptoms of scale formation are: (a) Consistently rising temperatures in the evaporator stages (decrease in vacuum). (b) Heating steam pressure to the feedwater heater, or heating tube bundle, is required to be above the design value t o produce the specified temperature at the feedwater heater outlet.

Sealed heat-transfer tubes can be cleaned chemically by circulating a diluted acid solution through, or over the tubes. Hydrochloric (muriatic) or sulfamic acid are the chemical reagents commonly recommended. Hydrochloric acid should be used only if sulfamic acid is not available, and then only by qualified personnel experienced in its use. Chemicals used for acid cleaning are samewhat haaardous; they are hazardous in the sense that most acids are dangemus. I t is essential that the operator understand the potential danger involved with the use of acids. If proper precautions are taken, persome1 injury and equipment damage can be avoided and acid cleaning of the evaporator can be accomplished in a minimum of down time. Sulfamic acid hasbecome the chemical mast commonly carried on board ship. No extra precautions are necessary in storing or handling dry sulfamic acid; however, since the acid dust will irritate the nose, eyes, and skin, careless handling which may result in its dispersion should be avoided. Sulfamic acid is only mildly corrosive to metals and there is no fire hazard involved in its use. lulfamic acid is considerably less objectionable in all respects than dilute hydrochloric acid. The maximum acid concentration should not be numerically lesa than pH 2.0. A reference for the correlation of pH values with various acidic and basic solutions is given by Table 4. When one pound of sulfamic acid powder with color

indicator is dissolved in one gallon of water, the solution will turn a light red color, indicating sufficient acid concentration to dissolve scale. If the solution is heated to approximately 120-140 F, the color will change to a deep red. As the solution is circulated through the evaporator components and comes in contact with scale deposits, a chemical reaction takes place between the scale and acid which dissolves the scale and reduces the concentration of the acid. The by-product of the chemical reaction is the liberation of large quantities of carbon dioxide gas which must be vented to atmosphere at some point in the cleaning circuit. As the acid is circulated and the chemical reaction takes place, scale is dissolved, consuming acid in the process, and the solution will gradually change in color to orange or yellow, indicating that most of the initial charge has been dissipated. At this time a recharge of sulfamic acid will be required to increase the acid strength and change the color back t o red. Beriodic recharging or makeup of fresh solution will be required until the acid solution remains red for to of an hour after the last acid charge. When this color condition occurs, the operator can be assured that all sohble scale in the cleaning circuit has been dissolved since the acid strength is no longer being dissipated by contact with scale. The acid should then be completely drained from the plant and the entire cleaning circuit flushed with large quantities of fresh seawater. 2.6 Distilling Plant Vacuum Equipmenf. Proper venting of the distilling plant condenser is most important to prevent the buildup of noncondensible gases. Air is liberated in the evaporator from the entering feedwater which is saturated with air. Carbon dioxide may also be released through a breakdown of bicarbonates in the seawater. Since a low-pressure marine evaporator o p erates under a vacuum, small quantities of air also leak into the unit through the gasketed joints. If the condenser is not properly designed, these gases collect in low-pressure pockets and render these areas ineffective, thereby reducing the performance of the unit. Figures 28 and 29 @howthe vent baffling and shrouding in a typical condenser and saltwater heater. Three different types of systems have been used to establish and maintain the low pressure required in the evaporation chamber; these are: steam-motivated air ejectors, mechanical vacuum pumps, and water-motivated air eductors. Of the three, steam-motivated air ejectors have been most common; however, mechanical vacuum pumps and water eductors are suitable for limited applications. The service conditions of a vacuum system vary depending on the design of the distilling plant. On a submerged-tube or spray-film distiIler, the shell temperature and vacuum are maintained at a predetermined value regardless of the seawater feed temperature, and the vacuum systems for these distillers operate at one suction condition. I n a flash evaporator, the shell temperature varies widely as the seawater feed temperature

Table 4

pH Value INCREASING ACIDITY

0 1.0 2.0

T@;l Purple Red

2.5

3.0 4.0 4.5

Pink Orange

5.0 6.0

Gold Yellow

NEUTRAL

7.0

White

INCREASING ALKALINITY

8.0 9.0 10.0

Correlation of pH with Various Acidic and Basic %lutions

Strepgth Very strong

pH of Industnd Chemicals Sulfuric acid 4.9% (1.ON) Hydrochloric acid 0.37% (LON)

13.0 14.0

Dark Brown Black

lO,oOo,aa, 1,000,000

Fairly strong Too weak to dissolve sde Mild Neutral

100,000

/

Orange juice Beer

A&;

acid 0.87 (0.1N)

10,000 1,000

n

100

American cheese M i

ib i

- .

Distilled water at 77 - .F Egg white

10

Borax Fairly strong

1w

Milk of Magnesia '

Ammonia 1.7% (1.0N)

12.0

Lodalue arithmic to Pure Water

Lemon juice

Will dissdve average scale

Mild Grass Green Dark Green

pH of Household Items

Caustic soda 0.4% (1.ON) Very strong

1,000 10,000

100,000 1,000,000 10,000,000

Caustic soda 4.0% {l.ON) NOTE: The color assumed by the indicator (e.g. litmus or pH paper) varies depending on the type and r a w of indicator used.

changes throughout the range of 28 F to 85 F. The vacuum system must, therefore, be designed to provide proper operation at the suction conditions encountered with inlet seawater feed temperatures in this range. In a two-stage flash evaporator, the temperature in the second-stage evaporation chamber will be approximately 114 F at a seawater temperature of 85 F; however, a t a seawater temperature of 28 F,?the second-stage temperature will drop to 70 F. The specific volume of steam is 238 cu ft/lb a t 114 F, but at 70 P the specific volume is 869 cu ft/lb. Consequently, the 28 F seawater feedwater condition controls the design of a flash evaporator vacuum system. The vacuum system must be capable of extracting large volumes per pound of steam and noncondensibles removed a t the lower seawater feed temperature. The highly corrosive nature of the noncondensibles released in a distilling plant vaporization chamber must also be considered in the design of a vacuum system. In addition to air inleakage and seawater deaeration loads, the distilling plant vacuum system must also remove the COz which results from the breakdown of carbonates in the seawater when it is heated to 170 F as well as NH;I and HpS, which are also introduced by polluted seawaters. a. Steam-Motivated Air Ejectors. Air ejector systems used with most marine evaporators are of the twostage noncondensing type, but singlestage systems can

be used on submerged-tube and spray-Mm units where higher shell pressures are maintained. The elementary operating principle of air ejectors may be understood by reference to Fig. 30. High-pressure motive steam is led to the unit and passes through the nozzle where its pressure is dissipated in accelerating the steam to a high velocity as it passes from the nozzle throat through t$e expanding section of the norele. The high-velocity jet of steam issuing from the nozzle entrains the saturated air mixture entering the ejector element. f i c t i o n between the steam jet and the low-pressure air causes the latter to move with the steam into the converging section of the diffuser tube where the steam and air mix. The divergent section at the downstream end of the difluser tube serves to decrease the velocity of the moving gas and increase its pressure, thus converting the kinetic energy to pressure energy. I n this way the air and noncondensible gases removed from the vacuum system are compressed, which is the object of,the air ejector element. Por a d a o n a l dis - fat,h.. - ~. . ~ .~. Weight of chain, lb. . . . . . . . . . . . . . Weight of anchor, lb. . . . . . . . . . . . . Tot,d weight = WI Wa,Ib. . . . . Hawsepipe efficiencyn.. . ... . . . . . . . Buoyancy factor.. . . . . . . . . . . . . . . . Static load at wildcat = e3bWa,lb. . Angle of brake wrap around drum, deg. . . . . . . . . . . .I . . . . . . . . . . . . . Angle of wrap = lr61/180, red. . . . . Brake lining coeff. of frictionb.. . . . Ratio factor = e w . . . . . . . . . . . . . . . Assumed velocity of fall, fps. . . . . . Specified stopping distance, ft. . . . . Deceleration force = e3W3V1/2h lb Force at wildcat = f i W4, fb!. . Force at brake drum = Fsd/D, lb. . Slack end r l l ->/(K - 1), lb. . . Anchor en pull - F3 Fa, l b . . . ? Handwheel effort = F4/2Lfelest lb. . Anchor end band stress = Fs/wtl, ps1 . . . . . . . . . . . . . . . . . . . . . . . . . . . Slack end band stress = F4/wts, psi. Mean band stress = (81 S2)/2, ps1 . . . . . . . . . . . . . . . . . . . . . . . . . . . Band modulus of elasticit , psi . . . Band stretch = ~ 3 ~ a / 2 l i n . .. . Handwheel turns to set braked = yBG1/L. . . . . . . . . . . . . . . . . . . . . . . Maximum 2Fs/wD, brake psi.. . .band . . . . .pressurea . . . . . . .. = .. ~-

~~

~

-

:'.

~ ~ - - -

Symbol

C G

P

a d D w ts

+

Fig. 17 Schematic of a band type of anchor windlass brake

1)

pansion as the binder breaks down and swells with heat. Molded linings are most effective under the conditions of the drop test if deeply scarified or if vented with lateral grooves about 3i6 in. wide by %6 in. deep, pitched about 23 in. around the circumference. A new brake should be "run in" by operating the windlass under power with the chain held clear of the wildcat. The band should be examined periodically and the operation continued until the area of the brake lining in contact with the drum surface is at least 75 percent of the total area of the brake lining. The drum and lining surfaces should be free of exudation; if need be, they should be cleaned with a solvent. Care should be taken that overheating does not occur during the run-in operation. These precautions should be accomplished at the shipyard rather than at the windlass manufacturer's plait. The new linings will then be in the best condition possible, arid preservatives and rust preventive media will be eliminated as prospective reasons for a reduction of the band lining coefficient of friction. The design of an anchor windlass brake is heavily influenced b y t h e experience gained with previous designs, as may be noted from the typical anchor windlass brake calculations shown by Table 3. The brake lining coefficient of friction used in the design calculations is generally somewhat less than the value suggested by the brake lining manufacturer and may a t first appear to be unreasonably low. However, the lining coefficient of friction suggested by the manufacturer is usually based on ideal laboratory conditions which hardly simulate anchor windlass brake service. Special instrumentation was provided during the sea trials for the aircraft carrier USS America (CVA 66) in order to confirm that the design criteria employed in the design of the-anchor windlass brake were adequate. Before the sea trial, every reasonable precaution was taken

+ +

1.0 0.5 H 20 2.0 do L 0.25 0.165 2 1 GI e1 0.92 M 503 f- - . 60 19,840 W Z 12,693 Wa 32,533 0.80 0.87 W4 22,600

tl

-

Type of Ship C4 CVA 66 2jis 4% 1498 28.5 19.0 9% 5 31 60.48 48 90 2 X 10 10 1.0 0.75 26 2.5 0.5 0.238 2.76 6 0.903 2705 60 80,000 60,000 140,000 0.80 0.87 97,440

"7

+

:

Notes for Table 3

An optimistically high hawsepipe efficiency should be used in this calculation. b. In general the brake lining coefficient of friction quoted by manufactur6rs should not be used, as that is the value ?btained?n a laboratory. Anchor windlass tests show that a coeffiaent of fr!ction of 0.225 and 0.30 can be expected with molded and woven lmings, respectively. c The braking force required at the handwheel should be approximately 100 lb or less. -. d Efforts to red~uce the ha&wheel effort by inaxeasing the mechanical advant.age may result in an excessive amount of elasticity in the system. The maximum pressure on the brake lining should be as recommended by the manufacturer and proven by experience. f In order to prompt1 initiate braking action, a hydraulic-assist mechanism was installeion the CVA 66. The mechanism wae des i r e d s~uchthat the brake was set hydraulically when the handw eel was turned about one-half turn in the direction to set the brake. Turning the brake handwheel also set and secured the brake mechanically as in conventional practice. ,,&+

HULL MACHINERY

MARINE ENGINEERING

COEFFICIENT OF FRICTION '

1

COEFFICIENT OF FRICTIW

d ,/CHAIN

VELOCITY

200-

160-

EN0 PULL

ANCHOR END PULL 0

0 '1.

U) W

100-

I

0 J 0

PULL

i

J

2

so-

TIME, SECONOS

(a) Drop to 15 fathoms 0TIME, .SECONDS

(c) 3 0 to 4 5 fathom drop OF FRICTION Z

Y

P

1

b0.4-

L L

0

COEFFICIENT OF FRICTION CHAIN VELOCITY

20.

0

; L

10.

0.2,

I*

t

Y

14. ANCHOR END PULL

SLACKENDPULL

10 TINE, SECONDS

(b) 15 to 3 0 fathom drop Fig. 18

USS America, CVA 66, anchor windlass brake test data

to ensure that the brake lining material was properly prepared and "run in." Figure 18 shows the results obtained during the CVA 66 tests, and the relatively low brake lining coefficient of friction anticipated (see Table 3) is seen to be confirmed.

A chain counter is a very useful tool that may be installed on a windlass. A chain counter provides a mechanical readout at the windlass of the number of feet or fathoms of chain that have been payed out, and provides an electrical readout in the wheelhouse. The officer

TIME, SECOND8

(d) 4 5 to 6 0 fathom drop Fig. 18 (continued)

586

HULL MA

MARINE EN(

on dutv then knows the amount of chain in use without sending a man to thc forecastle to check the markings on the chain. The indication of the amount oT chain out is also helpful (when the depth of water is known) in paying out enough chain to ensure that the anchor will hold. A mechanism may also be installed to provide a means for releasing the anchor or anchors from the wheelhouse. To set the windlass up for this operation, the hand brake on the wildcat is tightened, the wildcat clutch disengaged, and any devices used to secure the chain at sea, such as the devil's claw or tongue-type stopper, are removed from the chain. A hydraulic cylinder, powered by an accumulator, is arranged mechanically to override the brake screw mechanism. Then. when a solenoid valve is energized from the wheelhouse, the wildcat brake is released and the anchor will drop. Speed governing is built into the hydraulic system to limit the rate of fall of the anchor. This type of arrangement has been successfully used between a tug and an unmanned barge. I n this instance the control was accomplished by a radio signal, with the small amount of electric power required on the barge supplied by batteries. Remote control of the anchor would-be particularly advantageous in a long river passage where it is customary to have at least one man standing by at the windlass in case an emergency drop is necessary. d. Windlass Power Units. Since the late 1960's., verv " few steam-driven windlasses have been manufactured. However, steam-driven windlasses were common before then, particularly on tankers which carried inflammable cargoes. Steam-driven windlasses are usually of the horizontal type with all of the components located above deck; such a windlass would look much like the one illustrated by Fig. 13 if a steam engine were substituted for the electric motor. The steam engine is commonly a horizontal reversible type with two cylinders. Steamdriven windlasses are designed to operate with a steam pressure of about 100 psig at the throttle. Steam-driven anchor windlasses are inherently rugged, simple, and reliable; however, these considerations are seldom criteria for selection as other types of windlasses can be designed to be equally dependable. The major advantage associated with steam-driven windlasses is that they entail no fire hazards when used on tankers that carry inflammable cargoes. On the other hand, the long runs of piping from the engine room pose two problems: one of actually getting steam to the windlass, and the other of maintaining insulation on the pipes when they are run above the weather deck, as is the usual arrangement. These two problems are, obviously, closely related. In cold weather, if the windlass is steam-powered, it is usually necessary to turn the steam on well before the windlass will be needed so that steam and not condensate gets to the unit when it is needed for anchor handling. The two commonly used powering systems for windlasses are direct-connected electric motors and electrohydraulic systems. Electrohydraulic systems permit complete control over the hoisting speed and also provide

protection (by relief-valve action) against shock loadings in the t~.ansmissionshafting and gearing in the event that the anchor is inadvertently housed too abruptly. When an electric motor is directly connected to the windlass, it may be either a squirrel-cage or wound-rotor alternating-current motor or a direct-current motor. A d-c motor provides sufficient speed control to house the anchors safely. If a squirrel-cage motor is used, it should be of either the two- or three-speed type with the slowest speed usually one quarter of the full-load speed and slow enough to house the anchor satisfactorily. Even if a multi-speed a-c motor is used, the anchor should be driven through a slip-type clutch coupling so as to limit inertial loadings in the event that the anchor is housed too abruptly. If variable speeds are necessary or desired, then either a wound-rotor a-c motor or a d-c motor may be used. If a wound-rotor motor is used, it should be separately ventilated with the ventilation air taken from below deck so that water will not enter the air intake. The discharge air may be directed back to a protected space, or the air may be discharged directly on the open deck. In the latter case, the discharge outlet should be protected by a solenoid valve arranged so that the solenoid will open the air duct when the blower motor is energized. Interlocks must be provided so that the main motor cannot be energized unless both the blower motor and resistor fan motor are running; and additional interlocks should be provided to shut the entire system down in the event that the temperature of either the main motor discharge air or the resistor bank discharge air exceeds a predetermined safe limit,. If the variable speed for a motor-driven windlass is to be obtained from a d-c motor, the direct current may be provided by the adjustable voltage output from an a-c/ d-c motor-generator set. This drive is advantageous if the windlass is part of a combination unit, the other part of which is a constant-tension mooring winch. Windlass speeds and loads are controlled as described later under Cargo Winches and Constant-Tension R'Iooring Winches. In the electrohydraulic windlass, the pump, or A-end, is usually located below deck and driven by an a-c motor; and the hydraulic motor, or B-end, is mounted on the input shaft of the windlass gear reducer. When the windlass is of the horizontal-shaft type, the B-end ismounted in the weather. When the windlass is of the verticalshaft type, the B-end is mounted below deck out of the weather. Some vertical-shaft windlasses, partlicularly those on naval vessels, have two completely separate power plants, one for each wildcat or wildcat and capstan combination. These are then arranged so that, in the event of a casualty to one power unit, the other unit may be engaged (usually hydraulically) so as to operate both vertical shafts; see Fig. 16. A more common arrangement 011 merchant ships is one with a single, doubleended, electric motor driving two pumps with each pump discharging to a hydraulic motor. In each of the hydraulic arrangements described above, the hydraulic transmission would consist of a positivedisplacement, reversible-flow; variable-stroke pump

piped in a closed circuit to a fixed-stroke hydraulic motor. The first pinion in the gear reduction should be coupled to, rather than mounted on, the B-end output shaft. The hydraulic circuit should include an auxiliary, positive-displacement, replenishing pump. The pumping unit bedplate is usually built as a storage tank, and it should be large enough to contain 110 percent of the oil in the system so that all oil may be drained to the tank for servicing or maintenance. Some windlasses are provided with a horsepower-limiting device which is responsive to the pressure in the system. The horsepower limiter should be designed so that when the pressure reaches a predetermined value, the pump stroke is reduced with increasing pressures so that the electric motor horsepower remains constant. As the pressure in the system reduces (i.e., the anchor chain is hauled in), the horsepower limiter will return the pump stroke to the setting called for by the operator. Each side of the hydraulic circuit should be provided with a relief valve of the "cross-blow" type, returning the oil to the suction side of the pump. Alternatively, in some cases the discharge from the relief valves may be led to the sump tank. Direct electric-driven windlasses should be provided with an electric brake on the motor shaft. ,Hydraulic windlasses may be provided with either an electric brake or with a hydraulic brake which should be mounted on the B-end. Electric brakes should set upon loss of electric power, and hydraulic brakes should set upon loss of either electric power or hydraulic pressure. Speed and directional control for direct electricrdriven units should be effected by a master switch located aft and at a safe distance from the windlass. The master switch is usually equipped with a vertical handle, as described in Section 1. Stroke control for a hydraulic windlass may be effected either manually or by servo control. The control wheel stand should be raised to a convenient height above the deck and should be located aft and at a,safe distance from the windlass. If manual control is used, the shaft should be provided with a deck stuffing box through which it is led to the A-end stroking mechanism. The run of control shafting should be made with great care in order to avoid excessive friction, lost motion, and derangement that may be caused by deck deflection. Universal joints and slip couplings should be used to assist in eliminating the deleterious effects of misalignment. The handwheel for the hydraulic stroking device is usually arranged in a horizontal plane. An auxiliary handwheel for the stroking device should be located adjacent to the A-end to assist in servicing or warming up the unit. The stroke-control mechanism should be provided with a spring detent for the neutral position of the A-end stroking spiadle. Limit switches should be arranged to prevent the pump from being started if the pump and servo control are not in the neutral position. This is necessary to ensure that the anchor does not start to move when the pump is energized. Electric master switches should be provided with de-

Table 4 Power Calculations for an Electrohydraulic Windlass Type of Ship.. . . . . . . . . . . . . . . . . . . . . . . . . . . Number of ar~ehorshoisted. . . . . . . . . . . . . . . Anchor depth at beginning of hoist, fath.. . . Anchor weight, lb . . . . . . . . . . . . . . . . . . . . . . . . Anchor chain size, in.. . :. . . . . . . . . . . . . . . Chain weight (each anchor), lb. . . . . . . . . . . . Buoyancy factor.. . . . . . . . . . . . . . . . . . . . . . . . Weight per wildcat = b(W1 WZ),l b . . . . . Hawsep~peefficiedcya . . . .' . . . . . . . . . . . . . . . . Pull at each wildcat = WJ/e?, lb . . . . . . . . . . Outside length of one cham link, in.. . . . . . . . Pitch of links = G - 2C, in.. . . . . . . . . . . . . . Number of whelps on wildcat. . . . . . . . . . . . . Wildcat pitch radius = ap/?r, in.. . . . . . . . . . Torque at each wildcat = Plr, in.-lb. . . . . . . Electric motor speed, rpm. . . . . . . . . . . . . . . . Hydraulic pump speed, rpm.. . . . . . . . . . . . . . Hydraulic motor speed, rpm. . . . . . . . . . . . . . Gearing ratio: first reduction.. . . . . . . . . . . . second reduction. . . . . . . . . . third reduction.. . . . . . . . . . . Chain hoisting speed = 2?rrNa/l2R1RzRa,fpm Specified hoisting speed, fpm.. . . . . . . . . . . . . Hydraulic ump and motor efficiency. . . . . . Gearing efkciency: first reduction. . . . . . . . . second reduction. . . . . . third reduction.. . . . . . . Efficiency of wildcat. . . . . . . . . . . . . . . . . . . . . Torque per hydraulic motor = Tl/R1R&aezeae4eal in.-lb. . . . . . . . . . . . . . . . . Hydraulic pressure required,b psi. . . . . . . . . . . Total electric motor hp requiredc = nPIS/ 33,000 ele2ese4e5,hp. . . . . . . . . . . . . . . . . . . . . Electric motor hp provided,d hp. . . . .. . . . . . Capstan diameter, in.. . . . . . . . . . . . . . . . . . . . Capstan rope diameter, in.. . . . . . . ., . . . . . . .

,

+

Gearing, efficiency B-end = ezea. to capstan: . . . . . . . . . . . . . . . . . . .. . reduction = RlRz.. . . . . . . . . . . . . . . . . . . . Capstan rpm = Na/R4, rpm. . . . . . . . . . . . . . d)N4/12, fpm. . . . . . . Rope speed = r ( D Permissible rope pull = 33,000 ele$lz/S,, lb. .

+

Symbol n h W1 C WZ b Wa e. PI

G

P a r TI Nl NZ NJ Rl Rz Ra S Sl el eJ e4 e5 Tz

C4 2 30 12,693 0.87 19,673 0.60 32,788 14% 9% 5 15.5 508,200 1150 1150 1100 7.83 7.40 4.55 33.9 30 0.75 0.97 0.97 0.97 0.96

P,

2200 735

HI HZ D d

102.6 100 24 2.55

e6 R4 N4 S, F

0.94 57.9 19.0 132 17,625

Notes: a The hawsepipe efficiency used should be the lowest anticipated. b A check must also be made to confirm that there is sufficient capacity to hoist one anchor and the full scope of chain at no specified speed. In this arrangement, the electric motor has shafts on both ends, each of which drive a windlass through a Dension 60 hydraulic pump and motor. d Specifications often permit the electric motor to be overloaded 25 percent at the beginning of a hoist.

tents so that the operator can sense the speed position selected. Handwheels for controlling hydraulic windlasses should be equipped with a speed indicator, marked ''gll' ' ' ~ , '"~/4," ' and "Full," to each side of "Stop" for the neutral position. atio ion of %he handwheel in a cloc/kwise direction should start the windlass in the hoisting direction. The procedure follotved in determining the powering requirements of an electrohydraulic windlass with an attached capstan is shown by Table 4; the procedure for a direct-electric windlass is similar. Motors for direcb connected windlasses should have a 30-minute shorttime rating; whereas, for a hydraulically driven wind-

r,

4

1

II

588

MARINE ENGINEERING

HULL MACHINERY

jU SWIVEL FITTING FOR TOPPING LIFT

,--_ /

J(_SWIVEL

As shown by Fig. 19, the upper vang pendants, which are secured to the boom head, are used to swing the boom. Vang lines are required on both sides of the boom head. Lastly, the rigging arrangement must be capable of hoisting and lowering the load. The load would be secured to the cargo hook shown in Fig. 19 and would be hoisted and lowered by means of the cargo hauling part that goes to the cargo (hoist) winch drum. There are many variations of the basic rigging arrangement shown by Fig. 19. Reference [17] contains a discussion of some of the rigging arrangements which have been used and also relates the considerations which the naval architect must entertain when desighing cargo handling systems. Consequently, this chapter will concentrate on the design of the winches. Cargo winches are discussed in this section, and topping and vang winches are the subject of the following section. b. Mechanical. Ratings or duties of cargo winches have become standardized over the last 30 years. Most cargo winches are driven by 50-hp direct-current electric motors. Where hydraulic or steam winches have been installed, they also have approximately the same fullload speeds as the d-c motor-driven winches, but the light-line speeds may be different, being somewhat slower for the hydraulic units and somewhat faster for steam units. Cargo winches have different ratings at different line pulls; typical ratings for a 50-hp d-c motor-driven dlectric cargo winch are as follows:

PA0

Drum Line Pull, lb Drum Line Speed, fpm 0 500 3720 290 7450 220 14,500 105" 19,000 70' " "Two-speed" winch in low gear. Separate auxiliary drum.

OBLONG SWIVEL EYE

'

SHACKLE VANG PAD

Fig. 19

Nomenclature for cargo gear rigged for swinging or slewing

lass, the motor rating should be 30 minutes at 15 percent load followed by 30 minutes at full load. If a separate replenishing pump is provided, it should be rated for continuous duty. The main pump motor should be interlocked in such a way that it will not start until the replenishing pump is running. A start-stop pushbutton arrangement should be provided on or near the weather deck control station in addition to the buttons on the controller cabinet. A low-surface-temperature electric heater may be provided in the windlass sump tank to maintain the oil temperature at about 60 F in cold weather. This heater should be supplied from a circuit

independent of the main power feeders and arranged to be disconnected when the main pump motor is energized.

2.3 Cargo Winches a. General. The simple rigging arrangement shown in Fig. 19 has the capability of performing the three basic functions required of boom and winch cargo handling gear. One of the three functions required is the ability to top the boom, i.e., raise the boom head to the proper elevation. A second function required of the rigging arrangement is the ability to swing (or slew) the boom so as to control the transverse location of the boom head.

Most cargo winches are provided with a double gear reduction. The first reduction is frequently of the herringbone type while the second reduction is usually of the spur type. In some few instances, when the winches ate to be mounted on or adjacent to living quarters and very quiet winches are required, they are fitted with a single worm-gear reduction. Moat winches are of the singledrum type. If a gypsy head is desired on the winch, it is mounted on an extension of the drum shaft and the drum is then fitted with a mechanical brake and a clutch so that the drum may be secured when the gypsy is used. Occasionally a double-drum winch may be installed in place of two single-drum winches, thus saving the cost of the second motor and electric brake. On such a winch, both drums must be provided with clutches that are interlocked so that only one drum can be engaged at a time, but permitting both to be declutched at the same time. Also, each drum must be provided with a mechanical brake. A variation of the double-drum winch is a single-drum winch which drives an "auxiliary" or "derrick" drum through a third gear reduction. The shaft extension,

Fig. 2 0

50-hp, two-speed, d-c electric cargo winch without gypsy head

that might have had a gypsy head mounted on it, is fitted with a flexible coupling through which the third-reduction pinion is driven. The auxiliary drum will thus produce a line pull that is greater than that produced by the drum of the basic winch. Of course, the line speed on the auxiliary drum is reduced (see the foregoing winch rating data). Two of these winches are frequently used where a pair of burtoning booms are mounted on a pair of king posts and a heavy-lift boom is installed on the centerline. The hoist lines for the burtoning booms are led to the high-speed drums of the winches, and the hoist and topping lines for the heavy-li,ft=boomare led to the two auxiliary drums. Again, the clutches for the two drums must be arranged so that both can not be engaged at the same time. Still another type of cargo winch is the so-called "twospeed" winch such as shown by Fig. 20. In this winch, two pairs of second-reduction gears are provided, either of which may be clutched in to suit the load being handled. The output speeds of the alternative second reductions are usually in the ratio of about 2: 1. The output load and speed ratios may be varied to suit the rigging arrangement. The mechanical band brakes provided on cargo winches are intended to be holding devices rather than stopping devices. In an emergency, such as that caused by a power failure with a load suspended, the mechanical band brakes may be used to lower the load to the deck. For heavy-lift winches, the brake mechanism should be of the screw compressor type. On these brakes, a shaft with a handwheel has an acme thread which provides the force necessary to tighten the brake in a manner similar to the anchor windlass brake . a w n in Fig. 17. The brake bands are usually lined with a woven brake lhing similar to that used for the anchor windlass brake. In addition to the clutch interlocks mentioned above, it is advisable to interlock the brake and the drum clutch in such a way that the clutch cannot be moved out of engagement until the brake is set. As a substitute for the band brakes, it is frequently possible to use a simple locking device such as a bar or

MARiNE ENGINEERING

Fig. 21

Winch showing "double-diamond" type of spooling device

pin that can be inserted in a sleeve in the winch pedestal and pushed through to engage a hole in the flange of the winch drum. Drum clutches should be of the positive-engagement type; the most commonly used type is the jaw clutch. If the clutch is of the dry type, the jaws should be relieved a few degrees so as t o facilitate engagement. If the clutch is to be lubricated, then the jaw faces should be parallel, but a few degrees of backlash should be provided to facilitate the engagement. Brake and clutch linkages should be arranged and designed so that they will be extremely rigid in order that they will not be bent or otherwise deranged by being "forced." This is particularly true of clutch linkages. On many winches, there m a y b e no reason to declutch the drum for long periods of time. Due to disuse, it may then be necessary to use a large force to move the clutch and linkage, particularly when engaging the clutch. If the clutch linkage does not have sufficient stiffness, the position of theclutch shifter may not show the true position of the clutch jaws so that there may be only partial engagement of the clutch jaws. Insufficient stiffness in the clutch linkage has also been known to result in the clutch jaws working themselves completely out of engagement with the result that the load was dropped. (That is, even though the clutch handle was secured in the engaged position, the clutch linkage was so flexible that the clutch jaws separated and disengaged due t o internal clutch forces.) All clutches should be provided with a device that will secure the clutch handle in either the engaged or the disengaged position. The drums should be fitted with removable rope guards designed to prevent a slack rope from being wound over the flange and being wrapped around the drum shaft.

When a large amount of wire rope is to be stored on a winch drum, i t is advisable to use a spooling device. These devices ensure that the wire is distributed evenly across the drum and does not pile up in one place. Such a pileup may occur a t the center of the drum or against either flange depending on the lead of line from the heel block to the drum. The most elementary form of a spooling device is a grooved drum. The groove is a long spiral from end to end of the druni. However, a grooved drum will only control the storage of one layer of wire and only if the fleet angle is very small and if the wire is kept under tension continuously. Mechanical spooling devices may take any one of several forms, the most popular of which consists of a pair of rollers that are arranged with their axes a t right angles to the line lead and parallel to the drum flanges. The rollers are mounted on a trolley or carrier which is moved back and forth across the width of the drum to ensure that the wire is wound evenly across the full width. Translation of the carrier is provided by a doublediamond threaded shaft which is rotated by gearing that is powered from the drum shaft. A winch fitted with this type of spooling arrangement is shown in Fig. 21. The design of the gearing and the double-diamond must be such that the carrier moves one wire diameter for each turn of the drum. Another mechanical device is that known as the LeBus Fleet Angle Compensator, which relies on special drum grooving and a specially mounted fair lead sheave. There is additional information on this spooling problem in reference [17]. Openings should be provided under the-motors to permit inspection and maintenance of the underside of the motor and access to the studs that secure the field poles. c. Electrical. There are a number of ways of obtaining variable or multiple speeds when a winch is driven by an electric motor. One alternative is a single-speed squirrel-cage motor driving the winch through a variable &splacement hydraulic pump. Three- and four-speed squirrel-cage motors and wound-rotor slipring motors are also possibilities. The most popular type of drive on cargo ships is that which uses an a-c/d-c motor-generator set to produce adjustable-voltage direct current t o control the speed and direction of the winch. In the cases of the hydraulic and motor-generator set drives, the prime mover is started once during an entire cargo handling day; whereas in all other cases the prime mover must be started each time the winch is required to haul in or pay out rope. These frequent starts and stops will cause voltage fluctuations in the ship's system. The most popular system of control for cargo winches is the modified "Ward-LeonardJ' system, which employs an a-c/d-c motor-generator set to provide an adjustable voltage output to vary the speed of the d-c cargo winch motor. Usually a squirrel-cage a-c motor drives two d-c generators, each of which supplies a controlled voltage to one of a pair of motors driving two winches that are arranged for burtoning cargo. The motor-generator sets

HULL MACHINERY

Fig. 2 2

Vertical motor-generator set for cargo winches with rotor removed i

Fig. 23

Vertical motor-generator set to sene a pair of cargo winches

are located in deckhouses so that they may be of a dripproof construction. The control is usually of the solid- The lower part of the rotor in Fig. 22 is the a-c motor and state type and provides infinitely variable speeds both the two d-c generators are above. hoisting and lowering. The a-c motor should have a conA horigontal motor-generator set is illustrated by Fig. tinuous rating, but the generators and the d-c motors 24. The a-c motor in the center drives two d-c generamay have a 30-minute short time duty rating. The ch- tors, each of which powers one cargo winch motor. cuitry should provide safe lowering speeds if either a Multi-speed squirrel-cage motors with three or four brake failure or power failure, or both, occurs. The d-c speeds have been used very successfully for cargo winch motors may be shunt, stabilized-shunt, or compound drives. One such three-wed motm has 4, 8; and 28 wound, and of totally enclosed, nonventilated, water- poles; another uses 4, 8, and 36 poles; and% third has tight construction and should be provided with an auto- 2, 4, and 40 poles. I n each case, the high-speed (4 or 2 matic drainage fitting. pole) winding can handle loads up to about 3 tons drum Vertical motor-generator sets for cargo winches are pull a t about 300 fpm, and this is also the maximum shown in Figs. 22 and 23. The units are force ventilated light-hook speed. The second speed (8 or 4 pole) proby a fan on the rotor. Air is drawn in through the con- vides a line pull up t o about 6 tons a t about 150 fpm. troller door, passes through the base, and is discharged .An automatic step-back relay is provided so that if the vertically from the top of the motor-generator enclosure. line pull on the drum exceeds the nominal 3-ton rating,

592

MARINE ENGINEERING

Fig. 24

Horizontal motor-generator set

the motor will not accelerate to the high-speed winding even when the master switch is advanced to the highspeed position. The low-speed windings are used to land loads and to take up on a slack line. Although only three speeds are available, the transitions between speeds are quite smooth, the speed changes being absorbed by the inertia of the winch and the inherent elasticity of the entire rigging system. The four-speed motor offers 4, 8, 16, and 32-pole arrangements, with the 16-pole connection serving principally as a cushion between the slowest running speed and the third and full-load speeds. In each case, acceleration from one speed to the next is provided by time delays so that the motor will attain its speed before moving on to the next speed. An advantage that is obtained if squirrel-cage motors are used is that they are essentially trouble-free since there are no commutators, slip-rings, or brushes to be maintained. The control of the hoisting speed of wound-rotor slip ring motors is generally accomplished by inserting varying amounts of resistance in the rotor circuit. By inserting a large amount of resistance in the circuit, the starting torque of the motor is increased and the inrush (or starting) current is reduced. The smaller inrush current is beneficial in that there is less disruption of the ship's electric system each time a winch is started. As successive amounts of resistance are removed from the rotor circuit, the motor speed increases until, when all external resistance is removed, the motor will, in effect, be a squirrel-cage motor and will run at approximately synchronous speed. Although a rheostat could be used to provide the external resistance, it is more usual to remove the resistance in four or five steps, thus giving either five or six motor speeds. This method of control is essentially the. same as that used for the control of constant-potential

direct-current motor@supplied by a direct-current generator in the engine room. Both control systems are also similar in that the resistors produce large amounts of heat that must be~emovedfrom the winch equipment room by a forced-ventilation system. The control circuitry for the winch should be arranged so that the winch motor cannot be operated unless the ventilation fan is operating and also so that the entire winch system will shut down if the temperature in the equipment room rises above a safe limit. Special consideration must be given to the control of speed when lowering an overhauling load with a woundrotor motor. There are two common methods of doing this. One is by plugging or countertorque lowering, and the other is the unbalanced-stator method. In the countertorque method, although the master switch is moved in the lowering direction, relays sensitive to current caused by an overhauling load actually supply torque to the motor that opposes the load. Varying the amount of this counter torque changes the speed of lowering. When there is no load on the hook, the winch must be driven to pay out line. Since there is no load to produce current, the current-sensitive relays do not function and the motor drives the winch to pay out l i e at the maximum motor speed. In the unbalanced-stator method for controlling lowering speeds, unequal or unbalanced voltages are applied to the stator. Usually, an autotransformer is connected across one phase of the power supply. A series of taps from the transformer lead to contactors which furnish voltages both above and below normal. This results in applying two rotating fields to the motor, with one field rotating clockwise and the other counterclockwise. By changing the amount of unbalance, the relative strengths of the two fields change and the speeds are varied. There are several other methods of controlling both the hoisting and lowering speeds of wound-rotor motors, but these usually require elaborate circuitry and additional rotating equipment [18]. d. Hydraulic. Variable-delivery hydraulic pumps have been used to drive winches on a few cargo ships, and are also used on tankers where they replace steam winches which have their pipe insulation problems. These hydraulic units have been furnished in two ways: with the complete winch including electric motor, pump, and hf" draulic motor and brake on a common bedplate; or alternatively, with the electric motor and pump located in a deckhouse and connected by pipe to the hydraulic motor and brake mounted on the winch bedplate. With the latter arrangement, the electric motor usually drives two pumps, each of which serves one winch of a pair assigned to a pair of burtoning booms. With this arrangement, the capacity of the electric motor need only be about 150 percent of the capacity of the hydraulic motor to drive a single winch, since it is impossible to draw full load from both winches at the same time. Thus, the total installed electric horsepower is reduced. Because the motors and pumps are out of the weather, the motors can be dripproof.

I

593

HULL MACHINERY

a clean hydraulic system. Where the pumps are located in the deckhouse, the piping between the units must be fabricated and installed by the shipbuilders. Extreme care must be used to ensure the removal of dirt and other contaminants from the system by the use of a cleaning procedure such as is described in reference [I]. Special attention must also be paid to the provision of vent fittings at the high points in the hydraulic piping system for the removal of entrapped air.

Fig. 27

Worm-geared topping or vang winch fltted with disk-type electric brake

4

lowering the boom to the deck for servicing. In addition, topping winches are used to hoist the boom to the highest working position or, if the boom is to be stowed vertically, to the boom rest on the king post or crosstree. b. Mechanical. Topping and vaag winches are usually designed so that the lubrication and the fill-and-drain 2.4 Topping and Vang Winches connections will function properly whether the winches a. General. Topping and vang winches are used to are mounted horizontally or vertically. This allows the move the cargo boom vertically or laterally, respectively, shipbuilder latitude in arranging the winches so as to in order to position the boom head to handle the hook provide the best possible cargo handling arrangement. load (see Fig. 19). They are usually of simple mechanical The winches should be capable of handling the working and electrical arrangement, capable of positioning the and test loads without the aid of a drum ratchet and pawl boom when it is fully loaded, and are also capable of or other securing device. The drive arrangement may be

594

of either the spur gear (double or triple reduction) type as in Figs. 25 and 26 or of the worm-gear type as in Fig. 27. The 7.5-hp vang winch in Fig. 26 is designed such that the motor and brake are mounted within a vang post; only the drum is exposed to the weather. c. Electrical. Tcpping and vang winch motors are usually of the single-speed, reversible, squirrel-cage type, and are equipped with brakes. The motors should be rated for 30 minutes, short time, full-load duty. Master switches should be of the "spring-return-toOFF" type. Frequently the master switches for the topping and vang winches for a boom are arranged so that they are operated by a single lever called a "joy stick." As an example of the use of a joy stick, movement of the lever away from the operator will cause the topping winch to pay out wire and lower the boom; and movement of the lever to the right will cause the boom to swing to the right. A movement of the lever in any direction in between these two directions will both lower the boom and swing it to the right. Providing the control of the two winches in a single lever permits the operator to position the boom quickly and land the, load with great accuracy. 2.5

HULL M

MARINE EN(

1 DECK

Fig. 29

Capstans

a. General. There are three mechanical arrangements of capstans that are in general use. In one arrangement the motor, electric brake, gear reducer, and capstan head are mounted on a common bedplate on the weather deck. I n a second arrangement only the capstan head is mounted in the weather with the motor, electric brake, and gear reducer hung from the underside of the weather deck. I n a third arrangement the capstan head is on the weather deck with the motor, brake, and gear reducer on the deck below, as illustrated by Fig. 28. I n all three cases the master switch is located near the capstan head on the open deck, and the controller is located in a protected location such as a deckhouse or the capstan machinery space. The first arrangement mentioned has the advantage that the xomplete unit can be assembled by the capstan manufacturer for bolting in place by the shipbuilder. However, it has two disadvantages in that the motor and brake h u s t be of watertight construction and that the capstan head is elevated to an inconvenient height above the weather deck. The second arrangement has the advantage of having the motor and brake out of the weather so that they may be of dripproof construction. However, it introduces a deck penetration which must be made watertight, and the capstan head must be mounted by the shipbuilder. Neither of these problems is particularly difficult to handle; but the fact that the power unit is suspended from the overhead makes it more difficult to inspect and maintain. The third arrangement, which is common, entails a problem of alignment between the driving and driven units and requires the installation of a flexible coupling that can accommodate a small amount of misalignment

Warping winch with extended shaft

in Fig. 28 may be of either the roller type or of the sleeve type with bronze bushings. For merchant ships it is commonly specified that when the capstan is handling the specified load, the stresses should not exceed 40 per cent of the yield point of the materials. The capstan head, main shaft, bearings, and capstan base should be designed to withstand the breaking strength of the hawser applied tangentially a t midheight of the head without exceeding 75 percent of the yield point of the materials. c. Electrical. Capstan motors should be reversible constantand are usually of the two-speed (full and horsepower, squirrel-cage type. They should be rated for 30 minute short-time, full-load duty on either winding. A brake should be provided on the motor shaft. The motor and controls should provide adequate control of overhauling loads. Stepback protection from high to low speed should be provided so that, when retrieving a light line on the high-speed point, the motor speed will automatically step back to low speed if the rated horsepower is exceeded in high speed. Automatic return to high speed should also be provided should the line pull be reduced. Capstans are usually ,designed for line speeds of about 30 to 35 fpm in low speed since this is about as fast as a man can handle the line and keep it tight around the head for friction purposes. If the motor is full and quarter speed as suggested above, this will result in a light-line speed of from 120 to 140 fpm. 2.6 Warping Winches. A warping winch is typically used to warp a ship alongside a pier or to move a ship from one place to another, by means of hawsers, without other assistance. The warping head on a warping winch is similar to the head on a capstan, except that the warping head or heads are mounted on a horizontal shaft. I n some instances, the heads are mounted on extensions of the main shaft so that they may be at a considerable distance from the power unit, as in Fig. 29; if the shafts are extended in order to spread the heads apart, these extensions are provided with outboard bearing pedestals close to the heads. Occasionally a drum is installed, as in Fig. 30, for handling a wire rope for a stream anchor. If a drum is installed, it is provided with a clutch and a band brake of the screw compressor type so that the drum may

x),

n Fig. 28

DECK

Capstan w i h machinery on deck below

and variation in the vertical distance between decks which may be caused by either temperature changes or deck loadings. This type of capstan offers the advantage of having the power unit on the deck, rather than overhead, so that it is more readily serviced. b. Mechanical. The capstan head is usually of the smooth-barrel type (without whelps) and its configuration should approximate that given in reference [19]. The gear reducer generally consists of a worm-and-wheel reduction and a spur, helical, or herringbone reduction. It is preferred that the worm-and-wheel reduction (if used) precede the other reduction in order to take advantage of the slightly higher efficiency of the worm at the higher rubbing speed. All bearings in the reducer should be of the ball or roller type because of the necessiby of accurate alignment of the worm gearing. The deck bearings for the capstan arrangement shown

Fig. 30

Electrohydraulic mooring winch

be secured when the heads are being used. The electric motor, brake, and gear reducer are all mounted on a common bedplate. The warping heads should be of the smooth-barrel type (without whelps) and located a t a height that will ensure a suitable lead of line from the bulwark chocks. The gear reducer is usually of the double-reduction type with spur, helical, or herringbone g e a r i n ~but occasionally a single worm reduction is used. A stress analysis should be made for warping winches employing the limitations as described above for capstans; in addition, if a drum is provided, the unit should be capable of withstanding the stall torque of the motor without exceeding 75 percent of the yield point of the materials. The electrical equipment described for the capstans would be suitable for the warping winches. 2.7

Constant-Tension Mooring Winches

a. General. A constant-tension mooring winch is a device used to maintain a preset tension in a mooring line after a ship has been tied up a t a pier. The mooring lines holding the ship to the pier will maintain a constant tension even with changes in tide or ship's draft, without the necessity of manually adjusting the lines. When constant-tension mooring winches are used, two or more are installed forward and two or more are installed aft. As many as ten constant-tension mooring winches may be installed on ships such as&he largest$ankers. . Figure 31 illustrates a pair of constant-tension mooring winches that are rigged to lead the line from each over the same side of the ship through univer~alchocks. Alternatively, the lines may be taken to the other side of the ship. Occasionally each of two of the forward winches is combined with a wildcat to combine the windlassjconstanttension functions into a single unit. The constant-ten-

MARINE ENGINEERING

Fig. 31

Constant-tension mooring winches

sion winches are also arranged for manually hauling in and paying out the wire rope while the ship is being tied UP. b. Mechanical. The essential mechanical difference between a cargo winch and a constant-tension mooring winch is usually the interposition of a planetary gear set between the two usual gear reductions and the drum. In addition, a band clutch is provided with the clutch friction surface on the outer diameter of the ring gear on most constant-tension winches. When the band clutch is set, the ring gear is locked and the sun gear is then able to rotate the planetary gears, thereby rotating the spider which is directly connected to the winch drum. The band clutch is not primarily intended for "free spooling," which is a desirable capability if the mooring line has been secured to a bollard when there is a small amount of way on the ship. Instead, most control circuits contain a "drift" position which energizes only the brake, thus allowing the winch to overhaul, and thereby permits "free spooling" without the necessity of making manual adjustments at the winch. All bearings should be of the ball or roller type to minimize friction so that the tension-sensing mechanism is adequately sensitive to changes in line tension. A typical constant-tension winch has the ability to maintain, within a reasonable tolerance, a tension of from 8000 to 20,000 Ib. There are usually three or four intermediate settings between these two extremes. Tension is usually sensed by either a coil spring or a torsion bar. If the tension sensed exceeds the preset tolerance, the motor is energized to pay out or haul in line until the original tension has been reestablished. The band clutch is usually capable of holding from 125 to 150 percent of the maximum tension setting. For the winch cited above, the band clutch would be capable of resisting a pull of from 25,000 to 30,000 Ib. If a sudden surge were to occur, the clutch would slip thus preventing excessive forces from being developed within the winch. The tension-sensing mechanism and the band clutch and their linkages should be totally enclosed. Where

.

-

this is not possible, nonferrous or stainless steel parts should be provided. If a gypsy or warping head is required, it must bk arranged so that it will function with the drum declutched and secured. If the head is to be used for warping duty, then the head, shaft, bearings, etc., should be designed to withstand the breaking strength of the hawser applied tangentially at mid-length without exceeding 75 percent of the yield point of the materials. c. Electrical. Most constant-tension winches are powered by an adjustable-voltage d-c drive similar to that described under Cargo Winches. Other satisfactory drives include wound-rotor slip-ring motors and variable-delivery reversible-flow hydraulic transmissions, both of which are also described under Cargo Winches. Details of motors and generators are about the same as for cargo winches except that the motor should have a duty rating capable of sustaining continuous cyclical operation in a five-second total cycle. Thermal protection should be provided in the motor, and the circuitry should be arranged so as to shut down the motor and sound an alarm if the motor temperature exceeds allowable limits. One master switch should be provided for each winch at each side of the ship. Generally the two or three switches required on each side at the bow or stern are combined as duplex or triplex units and may be arranged for either bulkhead or pedestal mounting. The master switch to be used at any time may be selected by a selector switch mounted on each winch. Alternatively, each port and starboard pair may be arranged electrically so that if one is moved away from the OFF position, the other switch is disconnected. Each master switch should provide the following operating positions: fast heave, slow heave, automatic, off, drift, medium payout, and fast payout. When the master switch is set at the "automatic" position, the line pull will be dictated by the setting of the tension switch on the winch. d. Variations. There are three major variations or adaptations of the constant-tension mooring winch. These are: (1) the winch which has a warping head and can be used as a warping winch; (2) the winch that is combined with a wildcat so that it can handle a bower anchor and its chain; and (3) the winch that is designed to handle a synthetic mooring line on the drum. Constant-tension mooring winches are often fitted with "gypsy" heads of small diameter for handling miscellaneous lines. These are not practical for handling mooring lines of the usual 8 to 9 in., circumference. If it is desired that the winch be capable of warping the vessel, using such lines, heads approximating the sizes quoted in reference [19] should be fitted. These must be operable separately from the drum. A secondary shaft and gear reduction may also be required. The second variation is the constant-tension winch that, through an additional gear reduction, drives a wildcat that handles the bower anchor. This is an advantangeous arrangement since it is unlikely that it will be

HULL M

Fig. 33

Bow thruster

below deck, out of the weather, provided adequate fairleads are installed and a spooling device is mounted on the storage drum. 2.8 Fig. 32

I

I

I

1

Ship with bow L ~ S t e r

necessary to handle both the mooring line from the drum and the anchor chain at the same time. Some flexibility is permitted in the arrangement of winch leads and chain leads; one arrangement is such that the wildcat which handles the starboard anchor is on the port constanttension winch and vice versa. This is illustrated by Fig. 14. The possibility of locating the wildcats and the upper ends of the hawsepipes almost at will permits proper chain leads while still ensuring that the lead from the drum is adequate for all mooring requirements. The third variation mentioned above permits the use of a synthetic line as a constant-tension mooring line. There is considerable advantage to be gained from this arrangement since the line itself has much greater elas ticity than a steel line and therefore can absorb most minor adjustments by stretching or relaxing without the winch being energized. A new factor is introduced, however, in that synthetic lines under tension must not be stowed on a drum unless the drum is properly reinforced. In at least one known case, a synthetic line was stored overnight on a drum after being hauled in under "normal" (say 20,000 lb) line pull; the result was that the line exerted enough compressive force on the winch drum to crush it. In order to avoid such a situation, the winch drum may be arranged to act as a capstan, with only enough t u n s of line on the drum to ensure adequate friction. The line is then pulled off the drum and is wound on a second drum for storage purposes. This second drum may be located close to the winch or may be located

Special Thrust Devices

a. General. S~ecialthrust devices are ~rovidedon ships primarily to ifnprove their maneuverinicapabilities at zero, or substantially zero, ship speeds when the rudder is relatively ineffective. Theke are three general types of these special thrust devices. The most common type consists of a propeller that is installed in an athwartship tunnel; the propellers in these units may be of either the fixed pitch or reversible pitch type. Trainable thrusters, a second type, are commonly designed such that they can be lowered through the bottom of the ship and trained through 360 degrees so that thrust can be developed in any direction; however, there are other types of trainable thrusters. One type of trainable thruster is similar to a large outboard motor, and the active-rudder type is designed to be incorporated into the rudder. The activerudder type consists of a small shrouded propeller that is driven by a submersible electric motor built into the aft edge of the rudder. In a third type of thruster, a pump is qrranged to take suction from beneath the keel and discharge to either side, to develop thrust port or starboard as desired. There is a considerable amount of literature that deals with the subject of special thrust devices; however, the most comprehensive treatpent of thesubject is contained in-reference [20]. b. Tunnel Type. The tunnel type of thruster, which is widely known as a "bow thruster" when installed a t the bow of a ship, is illustrated by Figs. 32 and 33. "Stern thrusters," that is, similar devices installed at the stern of a ship, have been used but they are not common. The principal performance features of a typical series of bow thrusters are shown in Table 5. The performance

MARINE ENGINEERING

598

Table 5 Performance Characteristics of Typical Reversible-Pitch Bow Thrusters

Horsepower

Dheter

RPM

150

3'-7" 4'-3C 5'-5" 6'-7' 7'-11" 9'-2"

450 420

300

500

800

1200

1800

340 290 240 210

Thrust, Ib 4,500

7,900 13,200 20,400 30,200

44,100

HULL MACHINERY

20

Thrustlhp

30

26 26

16

25.5 25 24.5

0 0

el2

.I.

cn

of any particular unit may vary from a comparable one shown in Table 5. In general, an improvement in efficiency is obtained with a propulsor having a larger diameter and lower rpm. Bow thrusters are often designed such that the thrust developed is both variable and reversible. This is generally accomplished by using a constant-speed electric motor to drive a controllable- and reversible-pitch propeller. The unit is started with the propeller set at zero pitch; then, as the need arises, the pitch is adjusted so as to provide the desired thrust to either port or starboard. For the smaller bow thrusters, however, a fixedpitch propeller is sometimes used with a variable- and reversible-speed driver. When locating the tunnel in the ship, it is desirable to have the tunnel well forward so as to obtain the maximum turning moment from the thrust developed. The depth of the tunnel must also be considered, particularly "or vessels which will operate at light draft, because the thrust developed breaks down at shallow submergence depths. A test was conducted at the Philadelphia Naval Shipyard to investigate the effect of submergence depth on the thrust developed. The bow thruster tested was powered by a vertical 800-hp electric motor which drove a reversible-pitch propeller through right-angle bevel gears. The thruster tunnel diameter was 6 f t 7 in. The ship draft was varied and the bollard pull was measured when thrusting to port and to starboard. The wind was steady at only a few knots, and there was ample clearance both below the keel and to adjacent piers. Figure 34 shows the data obtained when thrusting to starboard; values when thrusting to port were similar. I t may be noted from Fig. 34 that when the top of the tunnel is submerged less than about 1.5 f t , there is a marked loss of thrust. The design of the junction of the tunnel and hull is another factor which "requires study. The tunnel openings definitely affect the resistance of the hull; however, the effect is difficult to accurately quantify because it is of a relatively small magnitude. With a well designed tunnel arrangement, it would be reasonable to expect an increase in the ship's resistance of about 1 percent. When it is desirable to minimize the resistance added by the tunnel, flow studies should be used as means of engineering the shape of the tunnel fairing and also the orientation of protective bars, if used. Protective bars are usually mounted along the flow lines in the area of the tunnel openings. Usually they

-

m A

6

z

I

4

0

Fig. 34

3

1 0 -1 2 8UBUERGENCE OF TOP OF TUISmL, FT

Effect of tunnel submergence on thrust developed by a bow fhruster

have a slightly downward slope when going aft, as may be seen from Figs. 32 and 33. Protective bars of several different designs have been used. In some designs several of the bars are bolted in place so as to be portable and permit access to the mechanism and also to permit a removable blade to be unshipped. However, this practice has not been uniformly satisfactory because the bars that are bolted on have a tendency to come adrift. Consequently, a preferable design is one with the bars simply welded in place. If access is required at a later date, the bars are burned off and rewelded; it is often necessary to burn off the bars to accomplish major work irrespective of the design due to rigging problems. Access problems are alleviated with some designs by the provision of a propeller blade removal hatch in the tunnel immediately above the propeller. This hatch is large enough for a man to get through, and no access is necessary fiom the tunnel openings. I n this case, to remove a propeller blade, the ship is ballasted until the tunnel is out of the water and the blade is lifted vertically up into the ship through the hatch. The tunnel is generally made of mild steel and is welded into the hull. To minimize tip leakage losses, the clearance between the blade tips and the tunnel should preferably be no more than 0.25 in. However, in practice it is difficult to maintain the propeller and the tunnel concentric. The diiculty involves the welding practices during installation at the shipyard. After installation, it is not uncommon to find the tunnel to be no longer circular, thereby requiring the blade tips to be ground off. For seawater service all bolts, studs, nuts, and other fastenings should be of monel. Propeller blades are usu-

I

i

I

ally of stainless steel or nickel aluminum bronze, with the propeller hub of bronze and the pod struts made of steel. The propeller is driven through a right-angle gear drive (usually of the spiral-bevel type) that is contained within the pod assembly. The pinion shaft or inputghaft of the right-angle drive extends out of the pod and through the tunnel assembly. The arrangement can be designed such that the pinion shaft penetrates the tunnel assembly at any angle desired; however, the shaft is normally either vertical or horizontal. With a horizontal arrangement, the shaft normally goes directly to the prime mover without involving another right-angle drive. In the case of a vertical drive, the prime mover can be located many decks above. In a common arrangement a vertical electric motor drives the thruster input shaft through auxiliary shafting. If the prime mover is a type which cannot be oriented vertically (e.g., a diesel engine), it is then necessary to use a second rightangle gearbox. The second right-angle drive preferably should be of the spiral-bevel type supported by oil-lubricated, heavy-duty, antifriction bearings. Flexible couplings should be provided between the prime mover and the shafting. The hydraulic power unit is often mounted at least 10 ft above the load waterline so as to avoid the need for a separate gravity tank. This head pressure is necessary to ensure that an adequate pressure is maintained on the oil seals in the propeller hub so that seawater will not enter the pod if the seals should leak slightly. The unit should have a motor-driven pump mounted on a reservoir, complete with necessary piping, suction filter, pressure gages, relief valves, etc. The reservoir should have a capacity in gallons of at least 2.5 times the capacity of the pump in gpm and should be fitted with direct-reading level gages or sight glasses, fill, drain, and vent connections, and access covers large enough to permit the reservoir to be cleaned. The master control stand is located in the wheelhouse and is sometimes made a part of the wheelhouse console. When a reversible-pitch propeller is used, this stand should contain a single lever that will pneumatically, electrically, or electrohydraulically control the hydraulic blade positioning system so as to provide stepless pitch control from zero to maximum either to port or starboard. For fixed-pitch propeller installations, the propeller speed and direction of rotation should be controllable. Frequently, auxiliary control stands are located on each bridge wing and are connected electrically, hydraulically, or mechanically to the master control stand. In addition to the control stand or stands, a control and indicating panel should be provided in the wheelhouse. This panel should contain pushbuttons for starting and stopping the prime mover, the hydraulic system for pitch control, etc. It should also indicate the alignment of the system (e.g., propeller pitch indicator) and contain lights apd alarms for critical pressures and temperatures. If the bow thruster is to be used only for docking, the main motor need only have a one to two hour full-power rating. If, however, it will be used for extended periods

\

Fig. 35

Thrust device that can b q IoQered and rotated 360 d&reer to develop thrust in any direction

of time, e.g., as an assist in steering the ship at slow speeds during a long passage, then it should be rated for continuous full-load duty. All other motors for the unit should be rated for continuous full-load duty. The controller should incorporate low-voltage protection, and interlocks should be provided to ensure that the main motor cannot be started unless the pitch control system is in neutral, that lubricating oil and hydraulic control pressures are available, and that the thruster room ventilation fan is in operation (as applicable). c. Trainable Type. The trainable type of bow thruster is the most versatile of all since, when it is lowered'from the bottom of the ship, it may be trained in any direction to move the ship to port, starboard, ahead, or astern. It is rather unusable as a docking device, however, since it extends several feet below the keel and therefore is quite suscep5ile to d a ~ a g e . In this arrangement t?ik propeller and drive'shaft are usually lowered and raised as a unit, as generally indicated by Fig. 35. The mechanism which trains the propeller also moves up and down with the rest of the machinery. The propeller is mounted in a protective ring or Kort nozzle which is designed to improve the thrust from the propeller. A closure plate is welded to the bottom of the protective ring; therefore, when the

600

HULL MACHINERY

MARINE ENGINEERING

9 = the maximum design wave slope (in terms of a wave height h and a wave length A, 9 =

sin-' ~rh/A),deg

I

1

C FIN SHAFT

The stabilizing moment developed by the fin stabilizers

M, = NaL

-

WTER LEVEL

(6)

where

M, = fin stabilizing moment, ft-lb N = number of fins a = lever arm between the resultant lift force and the ship's center of gravity, ft L = lift force per fin, lb

( a ) CROSS SECTION OF A YON-ARTICULATE0

FIN STABILIZER (ALSO KWWN AS A SPAM TYPE OR SINQLY ALL-MOVABLE STABILIZER)

The lift force developed by each fin can be expressed in the conventional manner as: Fig. 36

Active fin stabilization sketch

where unit is retracted it must be aligned fore and aft. The closure plate provides a nominally smooth hull. One use for this type of thruster is for station keeping on a survey ship when it may be necessary for the ship to remain in position over a particular spot while bottom sainples or cores are being taken. One such installation has a fixed-pitch propeller driven by a constant-speed squirrel-cage motor through an eddy-current coupling. Varying the excitation produces a propeller speed of up to about 95 percent of the motor speed. The difference in speed between the motor and propeller shaft, or slip, results in the creation of heat which must be removed either by a ventilation system or by a seawater-cooled heat exchanger. Various interlocks are required for this type of unit so as to ensure that the space ventilation fan, the eddy-current ventilating fan or cooling water pump, etc., are operating before the unit is lowered and the main motor is started. An additional interlock is required to ensure that the unit is in exact fore-and-aft alignment before being housed so that the closure will fit into the hull opening. d. Jet Type. In this third type of thruster, a pump with large clearances takes suction from the bottom of the ship near the centerline and discharges it to either a port or starboard discharge pipe to develop lateral thrust. The flow may be directed by a "splitter" which can proportion the flow to provide varying amounts of flow to each side. Mechanically and electrically this is the simplest, but least efficient, of the three types described. 2.9

Active Fin Stabilizers

a. General. Active fin stabilizers are fin-type control surfaces which are usually located just above the turn of the bilge near amidships, port and starboard. In a seaway, the angle of attack of the fins is varied continuously by automatically controlled tilting gear so as to

produce heeling moments that reduce the ship's tendency to roll. A compilation of the characteristics of many active fin installations is contained in reference [21]. Active fin stabilizers require ship forward motion in order to develop lift, which increases as the ship speed squared. In practical terms, this speed dependency limits the application of active fin stabilizers to ship speeds above approximately 12 knots. Below that speed, the fin size tends to get too large, and other devices (e.g., anti-roll tanks) find application, as described in references [22] and [23]. b. Simplified Calculation of Stabilizing Moment and Number of Units. The determination of percentage of stabilization and the design of automatic controls require a complex analysis. However, the fin size and tiltinggear machinery characteristics and location can be determined adequately using a simplified approach. With the simplified approach, it is assumed that a regular beam seaway having a small surface wave slope and a wave period approximately equal to the ship's natural period can build up large roll angles. By designing fin stabilizers to counteract the wave slope heeling moments, it is possible to reduce the large roll angles. Figure 36 illustrates the concept for a ship with two symmetrical fins. The rolling moment induced by the seaway is expressed as follows: Mo = 2240 AGM sin 9

CL = a nondimensional lift coefficient (Typical values of CL would be approximately 1.1 for nonarticulated fins, and 1.4 for flapped fins. Figure 37 defines the two types of fin surfaces. Section 8 of reference [5] thoroughly discusses the procedure for computing CL.) p = water mass density, lb-seca/ft4 A = area of one fin, sq ft V = water speed used-in fin design, fps (using VK as the ship speed in knots, V = 1.69 VK)

f

#

(5) R

where

I

Mg = roll-induced moment, ft-lb A = ship displacement, tons

GM

= metacentric height (the distance between the

ship's center of gravity, G, and the ship's metacenter, M), f t

.

/

k FIN

LFLAP

I

t b ) CROSS SECTION OF A FUPPIO FIN STABILIZER (ALSO KNOWN AS AN hRTICULATE0 OR DOUBLY n u - MOVIBU (CT~BILIZER)

If the stabilizing moment computed from equation (6) Fig. 37 Common types of fin stabilizen is equal to or greater than the induced rolling moment calculated from equation (5), an effective stabilization system is considered to be achieved. The key to this simplified method is the somewhat arbitrary selection of the seaway wave slope capacity 9 and the associated ship 9 = 5 deg speed VK. Experience has shown that stabilizers should CL = 1.2 be designed for wave slopes of about 4 or 5 deg. Lower a = 56 percent of B values are reasonable for very large ships, and higher values for small ships, since small ships are more likely The values of CL can later be refined (e.g., Fig. 31 of reference [24] indicates that the CL for a flapped fin to be subjected to roll excitation in a given seaway. varies between 1.6 and 1.3 from speeds of 5 to 25 knots, The lowest ship speed, VK, at which a significant roll reduction is desired, is usually taken to be approximately and Sectjon 8 of reference [5] gives data for nonarticu50 percent to 80 percent of full-power speed. This im- lated control surfaces as a function of aspect ratio, plicitly recognizes that ships in a heavy seaway are sweep, and taper ratio.) The usual practice is to try N = 2 and N = 4, and then make a final decision as to the generally operated well below full power. number of fins. Next, combinations of nonarticulated I n order to obtain a first approximation of the fin area versus articulated and nq;stsetractabb versus retractable required, the following expression may be used: fins are investigated in order to establish the preferred arrangement. Some of the considerations involved are discussed below. c. Location and Type of Units. The simplest mawhere B is the ship's beam in feet and all other terms are chinery and most compact arrqgement are obtained as previously defined. Equation (8) can be derived by with nonretractable units. On the other hand, the smallcombining equations (5), (6), and (7) with the tentative est fin area is obtained with high-aspect-ratio fins (e.g., assumptions that fins with an outreach of at least twice the fore-and-aft

MARINE ENGINEERING

HULL MACHINERY

tially the control system measures the instantaneous heel coypling) ; or (2) accepting only limited roll angle residual ship roll motion, assumes a simplified equation stabilization in order to reduce high-frequency roll anguof roll, and orders fin angles at the proper phase relation- lar acceleration. ship to achieve fih lift and corrective righting moments. Most automatic fin controls are not designed to correct Robb [26] provides an introduction to the theory, Lewis steady heel from ship's loading or wind, but rather to [27] presents more advanced theory, and Bell [28], stabilize around the mean heeled angle. This is done to Chadwick 1221, and Conolly [25] provide specialized de- conserve the limited fin stabilizing moments for correctsign procedures. Thus, one set of simplified control ing the oscillatory roll variations. equations is given in reference [22] as: Another feature in many designs is the automatic angle or fin lift-limiter. This is a consequence of having 4 21,-4 4 = 7' - 7 fins which develop full rated lift several knots below fullShip: W2 W power ship speed. At the higher ship speeds, the lift and 4 4 (10) torque, which are proportional to the ship's speed Control: a = K14 K2 - K3 squared, might be excessive and not really needed. Lift W u2 is limited to safe values by a fin angle limiter which is Fin: 7 = K,U2a (11) either in the fin lift control circuit, if such is installed, or in the fin angle servo circuit, using as input the ship spied where the symbols have the following physical signifi- deduced from the main propulsion shaft rpm. cance :

+

+

+

Fig. 38

Nonretractable articulated fin unit, port side, looking forward

+

4

length) that are located in the ship a t positions of maximum beam and located at angles which slope downward going outboard such that the fins have maximum leverage. The solution for the least fin area practically always involves fins projecting beyond the hull block dimensions, thus requiring retractability or a compromise design with low-aspect-ratio fins. Additional considerations to be entertained when developingthe arrangements are: (1) in the rigged-out position, the fins should be a t least a few feet below the design heeled waterlines (so as to minimize cavitation), and not immediately upstream of important sea chests (such as for main condensers); and (2) the fin tilting gear should be in a compartment suitable for routine maintenance (e.g., an auxiliary machinery space), and preferably not adjoining prime sleeping quarters. Clearly these contain built-in codicts, and require judgement of the naval architect and marine engineer to reconcile. I n commercial practice, the compromises generally result in one or two retractable, articulated fins per side. The retractable feature permits locating the fins in the ship at positions of maximum beam, where there is favorable leverage. The articulated (doubly all-movable) fins require about 30 percent less planform area than nonarticulated fins and are therefore easier to retract. On destroyer escorts, the U. S. Navy has used nonretractable units, some with articulated and some with nonarticulated fins. In this application, the hull form sections are much more rounded than cargo ship hulls, and also sonar domes extend several feet below the keel. Each fin axis is thus sloped well downward going outboard, and the fins do not protrude beyond the maximum beam or the sonar dome navigational draft. In this type, illustrated by Fig. 38, the tilting gear, fin, and hull insert plate can be readily furnished as one unit with a factorysealed hydraulic system. The fin unit shown is installed with the fin sloped 55 deg downward going outboard, and it has a span of 8 ft and chord of 4 f t (main body 36 in.,

.

tail flap 12 in.). Each unit has a total weight of 19,000 lb. The hydraulic tilting gear has a capacity of 226,000 in.-lb and is powered by a 20-hp motor. d. Fin Tilting Gear. The fin shafts of virtually all successful installations are tilted by hydraulic power, which is well suited to the requirements for rapid reversal. The tilting gear is generally similar to an electrohydraulic steering gear, such as described earlier in this chapter. Both Rapson-slide and vane-type rotary actuator installations are used; they are actuated by variable-delivery pumps whose flow direction is ordered by the amplified automatic control signal. Some of the significant differences from steering gear practice are: (1) There is usually only one power unit per fin shaft, since stabilization is not as critical a function as steering, and even if one unit fails there is another on the other side of the ship; (2) Whereas steering gears are not required to perform high-speed hard-over to hard-over rudder maneuvers for extended periods, the tilting gear may be heavily loaded for many hours of continuous operation, hence requiring appropriate motor ratings and system coolers; (3) Where retractable fins are used, the stowing and rigging mechanisms, interlocks, and indicators have to be provided; (4) Fin angular rates are about ten times greater than rudder rates (in Appendix 111 of reference [25], the minimum time for "hard-over to hard-over" fin angles is derived as 0.22 times the natural period of roll) ; (5) the rapid reversal of fin angles requires added torque to overcome the mass inertia of the mechanism (some allusions to this calculation are contained in Appendix B of reference [21] and in the author's closure of reference [241). e. Automatic Controls for Fin Stabilizers. Unlike the situation for rudders, human control of fin angles is ineffective. This is because the roll response of a ship in a seaway is at such a high frequency that the fin angles must be varied continuously and rapidly. Automatic controls have proved to be reliable and efficient. Essen-

603

roll angle of ship natural frequency of roll Ts = a damping ratio, involving GM,appendages, and hull form 7 = effective wave slope 7' = 7 2f , = ~ disturbing moment 7 = stabilizing moment of fins, expressed in terms of equivalent wave slope U = ship speed Kl9K2,K3 = constants of the particular system a = fin angle of attack = w =

+

i-

Usually a computer study is made to assure the desired stabilizer performance throughout the spectrum of sea conditions. Some coefficients can be calculated; others are estimated based on specific model tests or specialized hydrodynamic literature, such as reference [27]. In a computer study, the gains and sensitivities can be varied to suit the particular installation. Almost all successful controls use the ship roll velocity as the principal input, plus some roll angle. A case can be made for roll acceleration input and sway input in some applications. Many of the servo designs order fin angle, and measure the actual fin angle to get the closed-loop feedback. In some designs, e.g., reference [24], fin lift is used rather than fin angle. In typical shipboard designs 4 is measured by a gyro or athwartship accelerometer device, B, is measured by a rate gyro or by differentiating the roll angle, and 4 is measured by angular accelerometers. Some studies have been made of feedahead devices utilizing pressure taps port ancl starboard to get wave slope inputs. These devices have not been practical because the real seaway is not the simple wave shown in Fig. 36. Instead, the wave form is the sum of very many waves of varying height, period, and direction, as described in reference [27]. In most cases, the automatic controls are designed simply to reduce the roll angle. Several variations, however, may be considered, such as: (1) stabilizing to the apparent vertical (for improved passenger comfort in sway-

References

1 ('Cleaning and Protecting Integral Piping for Oil Hydraulic Power Transmission Equipment ," Military Standard MIL-STD-419. 2 John Flodin, "Hull Machinery" in Marine Engineering, vol. 11,edited by Herbert Lee Seward, SNAME, 1944. 3 Frank C. Messaros, "Steering Gear and Deck Machinery" in Modern Marine Engineer's Manual, vol. 11,edited by Alan Osborne, Cornell Maritime Press, New York, 1943. 4 A. M. Nickerson, ~ r . , ' a n dV. A. Olson, "Steering Gears and Their Selection," Tram. SNAME, 1952. 5 Philip Mandel, "Ship Maneuvering and Control" in Principles of Naval Architecture, edited by John P. Comstock, SNAME, 1967. 6 Karl E. Schoenherr, "A Program for the Investigation of the Rudder-Torque Problem," Marine Technology, SNAME, July 1965. 7 "Standard Specification for Cargo Ship Construction," Maritime Administration. 8 "Steering Gears, Electro-Hydraulic, Marine," Military Specification MIL-S-17803. 9 Rules for the ClassiJication and Construction of Steel Vessels, American Bureau of Shipping, New York. 10 R. M. Luke and F. P. West, Jr., "An Integrated Steering System," New England Section, SNAME, 1960. 11 William H. Hunley, "Anchor, Mooring and Towing Arrangements; Steering" in Ship Design and Construction, edited by A. M. DIArcangelo,SNAME, 1969. 12 "Guide to the Design and Testing of Anchor Windlasses for Merchant C%ips," Technical and Research Bulletin No. 3-15, SNAME, 1964. 13 "Dimensions of Wildcats," Navy Department Drawing 52601-860304. 14 Catalog, Baldt Corporation, Chester, Pa. 15 A. M. Nickerson, Jr., and C. H. Russell, "Anchor Handling and the Anchor Windlass," SNAME, New England Section, 1957. 16 "Windlasses, Anchor, Electric-Hydraulic, Verti-

MARINE ENGINEERING

cal and Navy Type, Naval Shipboard," Militajr Specification MIL-W-19623. 17 D. F. MacNaught, "Cargo Handling" in Ship Dasigrl and Construction, edited by A. M. D'Arcangelo, SNAME, 1969. 18 P. B. Harwood, Control of Electric Motors, John Wlley & Sons, Inc., New York, 1952. 1g "Capstans and Gypsy Heads," Navy Department Drawing 82601-860303. 20 S. Hawkins, "The Use of Maneuvering Propulsion Devices," Robert Taggart, lnc., Report RT-8518, prepared under contract MA-3293 for the Office of Research and Development, Maritime Administration, January 21, 1965. 21 Hector T. McVey, "Anti-Roll Fin Stabilizers," Philadelphia Section, SNAME, 1959. 22 J. H. Chadwick, Jr., "On the Stabilization of Roll," Trans. SNAME, 1955.

1

23 A. J. Giddings and R. Wermter, ('A Survey of Ship Motion Stabilization," 5th Symposium on Naval Hydrodynamics, Office of Naval Research, ACR-112, 1964. 24 J. E. Flipse, "Stabilizer Performance on the SS Mariposa and SS Monterey," Trans. SNAME, 1957. 25 J. E. Conolly, "Rolling and Its Stabilisation by Active Fins," Quarterly Trans. Institution of Naval Architects, London, vol. 111, no. 1, January 1969. 26 A. M. Robb, Theory of Naval Architecture, Jlondon, 1952. 27 E. V. Lewis, "The Motion of Ships in Waves" in Principles of Naval Architecture, edited by John P. Comstock, SNAME, 1967. 28 J. Bell, "Ship Stabilization Controls and Computation," Trans. Institution of Naval Architects, London, vol. 99, 1957.

Burr

I Electric Plants Section 1 Introduction

1.1 Nature and Scope of Electric Plants. A- complete shipboard electric plant is similar to the electric power generating, distribution, and utilization system of a self-contained shore-based industrial installation. Electric power is required for motors driving propulsion plant auxiliaries and deck machinery, interior and exterior illumination, navigation lights, ventilation and air conditioning, stores and cargo refrigeration, electric heating, galley equipment, drinking water and sanitary systems, and casualty control machinery such as fire and bilge pumps. Power must also be supplied for interior communication systems, announcing and alarm systems, radio communication, radar, and other electronic aids to navigation. For passenger vessels, the electric power requirements extend to hotel and recreation loads, theater and dance floor lighting, restaurant and swimming pool equipment, motion picture projection, public address systems, and stewards call systems. For passenger and prew safety, the electric installation includes automatic fire detecting and alarm systems, power-operated watertight doors, and electrically released, spring-closed fire screen doors. Electric power is vital to all shipboard operations and to the safety and comfort of the passengers and crew. For this reason, shipboard electric plants must contain equipment necessary to maintain continuity'of service, since a vessel at sea is isolated from external sources of electrical energy. Therefore, standby ship service generating capacity, usually equal to the rating of one of the ship service generators, is provided. In addition, one or more sources of emergency power, designed to automatically assume load upon loss of ship service power, are required to supply those loads that are necessary for the of the passengers and crew; the source of power should also have additional capacity adequate to supply those loads vital to getting the pmpulsion plant and generators back in service. Quick-starting diesel generators are usually provided for emergency power; however, storage batteries or gas turbine driven generators are satisfactory for this service. Emergency - - storage batteries combined with d-c/a-c motor generator sets are required on passenger vessels to provide temporary emergency power to certain vital loads until the emergency generator can start and assume the entire emergency load.

To avoid prolonged shutdown at sea, adequate spare parts should be stowed aboard ship to replace vital parh which are subject to wear and breakdown. It follows that adequate detail drawings and manuals containing instructions for operation, repair, and adjustment also should be placed aboard ship. For greater dependability at sea, electric equipment necessary for the operation of the vessel is required to have certain marine features such as dependable operation during rolling and pitching of the vessel, mechanical parts resistant to shipboard vibration, and windings and hardware resistant to moisture and corrosion. A shipboard electric plant includes: generating equipment; switchgear for control of the generators and distribution of power; and distribution panels, transformers, motor generators, and bus transfer equipment as necessary to provide the proper type of power to electrical loads. 1.2 Rules and Regulations. Merchant marine electrical installations must comply with a number of laws promulgated by the United States Government. The following is a listing of such rules and regulations. 1 United States Coast Guard, Code of Federal Regulations, Title 46-Shipping. 2 United States Coast Guard-Rules of the Road, International-Inland. 3 United States Department of Health, Education and Welfare-Standards of Sanitation and Ratproofing for the Construction of Vessels. 4 Federal Communication Commission Rules and Regulations. I

It should be noted that the requirements of the Intemationd Conference of Safety of Life at seaof 1960 are incorporated in item above. addition to the rules and reguhtions imposed by law,

11

d

compliance with other rulesfG usually required by the specifications for building particular ships. The most generally included in ship,s specifications am as follows: 1 American Bureau of Shipping Rules for Building and Classing Steel Vessels, Lloyd's Register of Shipping Rules and Regulations for the Construction and Classing of Steel Vessels, or similar

I I

11

606

MARINE ENGINEERING

classification society rules depending 'on which society's rules are selected as a prerequisite for marine insurance coverage. 2 IEEE Standard No. 45-Recommended Practice for Electric Installations on Shipboard, published by the Institute of Electrical and Electronics Engineers. 3 Suez Canal Authority Rules of Navigation. 4 Illuminating Engineering Society-Marine Lighting. 5 Marine-Type Electric Lighting Fixtures UL 595 published by Underwriters Laboratories, Inc. 1.3 Alternating-Current Electric Plants. Alternatingcurrent plants are the standard for most marine installations. Selection of a-c plants as standard over d-c plants -provides many significant advantages, e.g., savings realized in first cost, weight, and space requirements, reduction in maintenance effort, better availability of equipment, and increased reliability. Many of these advantages are realized through the use of squirrel cage motors in lieu of d-c motors having commutators and associated brush rigging. A frequency of 60 cps is recognized as the standard for a-c plants. Alternating-current plants may be any one of the following types: 120 volt, 3 phase, 3 wire 230 volt, 3 phase, 3 wire 450 volt, 3 phase, 3 (or 4) wire The 120-volt, 3-phase, 3-wire a-c generating plant with 115-volt, 3-phase light and power distribution is suitable for only small vessels having few motor-driven auxiliaries. The 230-volt, 3-phase, 3-wire acc generating plant with 220-volt, 3-phase power distribution and 115-volt, 3-phase lighting distribution through transformers is acceptable to the regulatory agencies. As an alternative, a 120/20%volt, 3-phase, 4-wire system may be used for both power and lighting on small ships without the necessity of using transformers. However, these plants are seldom selected since there are no advantages to be obtained when' compared with the standard 450-volt, 3-phase, 3-wire plant. The 450-volt, 3-phase, &wire a-c generating plant with 440-volt, 3-phase power distribution and 115-volt, 3-phase lighting distribution through transformers is the standard for most marine installations. However, the 450-volt, 3-phase, 4-wire, a c generating plant provides advantages and should be considered for' vessels having extensive 208- or 220-volt container refrigeration requirements. On vessels with'a-c electric propulsion machinery, the high-voltage main propulsion generator may be used to power large motors driving cargo handling machinery when the vessel is in port. Cargo pumps on electrically propelled tankers are typical high-voltage large-motor applications. Usually, 2500 kw is considered the maximum feasible capacity for 450-volt, 3-phase shipboard generators.

Larger capacity generators a t this voltage will not generally be used because of generator design limitations. As the demand for electric power aboard ships increases, particularly for nuclear propulsion plants and vessels with unusually large cargo handling power requirements, generators having a capacity larger than 2500 kw will be required. To provide this increased capacity, it is feasible to install prime generating plants designed for 2400- or 4160-volt, 3-phase output. In this type of installation, large motors would be powered from the high-voltage system with normal ship service loads of 450 volts supplied from the high-voltage system through transformers. 1.4 Direct-Current Electric Plants. Direckurrent plants are economically feasible on relatively small vessels and on those intermediate-size vessels on which a preponderance of deck machinery requiring d-c wide speed range motors and controls is installed. Directcurrent generating plants may be of either of the following types: 120 volt, 2 wire 240/120 volt, 3 wire The 120-volt, 2-wire d-c generating plant with 115volt light and power distribution is suitable for small vessels having few motor-driven auxiliaries. The generators are usually rated a t 75 kw or less. The 240/120-volt, 3-wire d-c generating plant is the most usual arrangement used on d-c vessels. This arrangement provides for 230-volt, 2-wire distribution to power loads and 230/115-volt, 3-wire distribution to lighting panels. The neutral of this 3-wire dual voltage system should be solidly grounded. Lighting fixtures, appliances, bracket fans, etc., supplied from lighting panels are connected between positive and neutral lines and neutral and negative lines in as nearly as practicable equal loads. 1.5 Drawings. The following is a listing of the types of electrical system drawings usually prepared by the shipbuilder. For a detailed listing of plans required for a specific ship, reference may be made to the detailed specifications for building a similar ship. 1 One-line diagram for power distribution system. 2 Isometric wiring diagrams for power system feeders and mains. 3 Isometric wiring diagrams for lighting system feeders. 4 Elementary and isometric w i h g diagrams for interior communication and electronic systems. 5 Deck arrangement plans for power, lighting, interior communication, and electronic systems. 6 Power system list of feeders and mains. 7 Lighting system list of feeders and mains. 8 List of motors and controllers. 9 Wireway locations and details. 10 Electric load analysis of generating plant. 11 Fault-current analysis and voltage dip calculations. 12 Application and coordiaation of protective devices. 13 Switchboard schematic and instrumentation.

ELECTRIC PLANTS

Generating Plants I

I

1:

.,

.

2.1 Load Analysis. To determine the correct aggregate rating for a generating plant, it is necessary to establish the probable peak loads under the various operating conditions of the ship. This is accomplished by the preparation of an "Electric Load Analysis," an example of which is shown in Table 1. This analysis is a detailed tabulation of the total connected load and the operating loads at sea, during maneuvering and a t port. Operating loads are determined by applying a service factor to the expected connected load for each application for each operating condition. The service factor assigned to each application is a combined load factor and diversity factor representing the percent of its own possible maximum that is contributed to the load on the generator plant over a %-hour period. Ocoasional loads such as fire pumps, anchor windlass, capstan, and boat winches are assumed to have a zero factor under all operating conditions. 2.2 Aggregate Generating Capacity. The aggregate generating capacity, exclusive of any, emergency and propulsion generating equipment, will always be greater than the peak load determined by the load analysis. Ship generating plants must consist of at least two ship's service generating sets so rated that, with one set not in operation, the remaining set or sets can carry the necessary sea load under normal operating conditions without exceeding the normal rating. The probability of installing additional electrical loads a t some future time should also be considered when determining the aggregate generating capacity. 2.3 Number and Rating. After determination of the peak load, the next step is the selection of the most desirable number and individual rating of the generating sets. The factors to be considered are: (a) the first cost, (b) operating cost, (c) size and weight, and (d) desirability of using generators of a standard size. Figure 1 (for turbine-driven generators) and Fig. 2 (for diesel-driven generators) show the variation of factors (a), (b), and (c) plotted against generator rating. No precise mathematical evaluation of the optimum number and rating of generator sets is possible; therefore the final selection must be based on experience, consideration of data delineated by Figs. 1 and 2, and available standard generator ratings. For peak loads up to about 2200 kw, two generators are usually provided, each capable of carrying the total sea load. There is a trend with steamships to fit one turbine-driven and one diesel-driven generator usually of equal capacity. This permits the full sea load to be handled in the event of a boiler failure. Suitable arrangemerits can be made for parallel operation of the turbogenerator and the diesel generator; such an arrangement does not preclude the fitting of an emergency generator. For loads greater than 2200 kw, three or more generators are usually installed with the total number of

generators selected providing maximum flexibility, reliability, and continuity of service for the electrical distribution system. 2.4 Location and Installation. On most vessels the ship service generators are located in the main engine room. This reduces the number of watch standers and provides the most economical piping arrangement since piping for generator sets and associated auxiliaries and for propelling machinery are comporlent parts of the same piping systems. On large vessels with two or more engine rooms, there is usually ope ship. service generating plant complete with auxiliaries and switchboard in each of the engine rooms. In general, ship service generators and associated switchboards are normally arranged so that the generators are in view of the switchboard attendant and so that the run of electric cables from each generator to the switchboard is as short as possible. Emergency generator sets must be located ~ b o v ethe freeboard deck, aft of the collision bulkhead, and outside the machinery casing to satisfy regulatory body requirements. All generating sets should be installed with the shaft fore-and-aft so that rolling of the vessel will not cause undue loads on the bearings due to gyroscopic effects,or cause oil to spill from the bearing bousings. 2.5 Generator Set Characteristics. The considerations governing the choice of a d-c or a-c electric plant and the selection of a distribution system and the voltage a t which it is to operate are discussed in Section 1. These basic features having bee^ determined and the number and rating of generating sets determined, the characteristics to be specified by the ship design engineers in the request for bids from manufacturers aye discussed in the following. The prime mover for ship service generators may be either a steam turbine, gas turbine, diesel engine, or a combination thereof, Emergency generators are usually diesel driven. A discussion of the different types of turbines and engines may be found in Chapters 5, 6, and 7. The generating set is selected after an evaluation of bids has been made, taking into consideration the price, weight, dimensions, fuel consumption, reputation of engine and generator builders, and the quantity of similar sets already in service. Important factors in the selection are simplicity, rdiability, and maintainability. a. Steam Turbine Prime Movers. Turbine-generator sets designed especially for installation, operation, and servicing on shipboard are available. They are designed for single or parallel operation for long periods without shutdown for maintenance or repair. Marine generator turbines are horizontal, multistage, and of the axial-flow impulse type. Each marine turbine generator set should be compact and complete with turbine, speed reduction gear,

, -"

1

I

I

i

I, I

1 1

I1

11

li I,

1

B 1

I II 1 I

I

I

610

ELECTRIC PLANTS

MARINE ENGINEERING

STEAM RATE AT FULL LOAD 3

9

,

8

0 I

23

C I-

a

2

3 I-

Y

3

w

-

ul

z . 5 9 2.0. k.40 P

240

7

6 5

2

5.10 z3 1 0 3 2 0- 9- 0",

0 -

W

u

Fig. 1 Approximate unit cost, weight, and deck area required and steam rate of geared-steam turbine generator sets

4

h.30 z 3 0 5 k.20 :20 2

0

3

Fig. 2 Approximate unit cost, weight, and deck area required and fuel consumption of diesel-driven generator sets

200 300 400 500 R A T E D KILOWATT OUTPUT

l

I

0

500 1000 1500 2000 R A T E D KILOWATT OUTPUT

NOTE: DATA IS B A S E D ON 1800 RPM, DIRECT- CONNECTED

2500

FOLLOWING CHARACTERISTICS: 1200 R P M , 450 VOLT, 3 P H A S E , 60 C Y C L E A C G E N E R A T O R ; S T E A M CONDITIONS 850 PSIG 950 D E G R E E S F. T O T A L T E M P E R A T U R E . A N D 28.5 INCHES OF M E R C U R Y '

generator, rotating exciter (if used), condenser, condenser air ejector, gage board, and a self-contained lubricating oil system consisting of a cooler, reservoir, strainer, attached pump, hand-operated pump, and electric motor driven pump (when specified), all supported on a common base with provision for lowering into place as a unit. For some units the shipbuilder may prefer to procure or build the condenser separately and combine it with the set a t the building yard. The set should be complete with all attached piping and accessories supported in place and ready for connecting to shipboard piping. Pipe connections should be either flanged, brazed, or welded. The design should preclude leakage of vapor, water, and oil to the outside and accumulation of liquid in pockets. The turbine should be complete with the following equipment. 1 Steam governing valve assembly. 3 Oil relay type constant-speed governor or electrohydraulic load-sensing speed governor when small speed variations and quick response are desired. 3 Speed adjusting (synchronizing) device with local manual adjustment and electric motor or potentiometer for remote adjustment from the switchboard when synchronizing generators. 4 Combined trip and throttle valve held open by lubricating oil pressure and automatically shutting off steam to the turbine upon loss of oil pressure; also for admitting steam gradually by hand when starting up a set. 5 Interlock switch for energizing the generator circuit-breaker tripping device, to disconnect the generator from the switchboard in case the throttle,trips closed. 6 Overspeed governor independent of the constant-

speed governor, which upon overspeed of not over 15 percent will relieve oil pressure from the throttle trip valve to shut down the turbine. 7 Steam sealing manifold for pressurizing the turbine shaft packing to prevent entrance of air into the turbine, also piping and valves for drainage of steam leakage through the packing. 8 Automatic atmospheric relief valve for exhausting to the atmosphere in case a high exhaust back pressure occurs in the turbine casing due to a condenser malfunction. 9 Sentinel valve for sounding an alarm before the automatic atmospheric relief valve functions. 10 High exhaust back pressure trip device which will relieve oil pressure from the throttle trip valve to shut down the turbine. 11 Manually operated rotor-turning device for use during maintenance. 12 Reduction gear, generally single reduction and single helical, with pinion and low-speed gear shafts supported by two bearings. The low-speed shaft is generally flexibly coupled to the generatq shaft. 13 Steam supply strainer for protecting the turbine. The strainer may be integral with the combined trip and throttle valve. Marine turbine-generator sets are provided with a self-contained lubricating oil system. A gear-driven oil pump supplies oil to the turbine, reduction gear, and generator bearings, and also to the constant-speed governor, trip throttle valve, high exhaust back pressure trip device, and overspeed governor. The pump takes suction from the reservoir and discharges through a magnetic duplex-type strainer and a tube-type cooler. Oil coolers are generally designed to use 85 F (maximum)

450 VOLT, 3 P H A S E , 60 C Y C L E A.C. GENERATORS. S E T S A R E SELF,- C O N T A I N E D W l T H B A T T E R Y START!NG EQUIPMENT: C U R V E S A R E FOR S E T S W l T H 2 C Y C L E ENGINES.. F U E L CONSUMPTION A N D P R I C E OF 4 CYCLE E N G I N E S W l L L BE APPROXIMATELY T H E SAME BUT W E I G H T A N D DECK A R E A W l L L BE S L I G H T L Y HIGHER. PRICE, WEIGHT, A N D DECK A R E A O F LOWER S P E E D E N G I N E S W I L L B E A P P R E C I A B L Y H I G H E R A N D FUEL CONSUMPTION W l L L BE SLIGHTLY H I G H E R .

seawater for cooling and are provided with zinc anodes exhaust from auxiliaries when the ship is a t anchor. on the seawater side to minimize corrosion. The coolingDue to the trend toward automated vessels, additional water pressure should be maintained lower than the oil monitoring and control devices such as those devices pressure to avoid seawater contamination of the oil in necessary for remote start-up, operation, and shutdown the event of a cooler tube failure. A hand-operated of the turbine-generator set may also be incorporated in lubricating oil pump is provided for use during start-up the design of the set. and maintenance. Generally, the lubricating system b. Diesel Engine Prime Movers. Diesel engine also includes an electric motor driven pump which is driven ship service generators should, like turbine started automatically by a pressure switch upon failure generator sets, be designed for continuous operation for of the turbinedriven pump; this pump may also be used long periods of time, singly or in parallel. They are during start-up and shutdown operations. When spec- usually located in the engine room. Because of their ified, the lubricating gystem contains switches for low relatively slow normal speed, diesel-generator sets are oil pressure and high oil temperature alprms. The oil appreciably larger and heavier than turbine-generator level in the reservoir is measured by a dipstick and a sets. Sets --rated rqt_ 150-kw and below are available a t visible level indicator, when specified. S J X up ~ ~tcJ800 rpm. Above 150 kw, most sets are Gageboards should include, as a minimum, pressure d e s i g n e n r either 900 or 1200 rpm. Also because of gages for the inlet steam, gland sealing steam, bearing the low speed, the generator is coupled directly to the oil, and oil pump and thermometers for oil to and from engine. The generator may be of the two-bearing type the cooler. with the rotor flexibly coupled to the engine crankshaft The gageboard and control devices such as speed or may have, a front bearing only with the rear shaft end adjustment knob, throttle trip valve handwheel, hand oil coupled rigidly to, and supported by, the diesel engine pump handle, gland seal steam control valve, hand crankshaft. Generator bearings may be integral with shutdown trip lever (or button), tachometer, and valve the generator end brackets or supported separately by for regulating the oil-cooler water should be located so pedestals. If a rotating exciter is used, the armature that one man can start, operate, and secure the set. will be overhung on the gen-tor front- haft extension. A motor-driven condensate pump and a condenser Two- and four-cycle engines are available for marine circulating water pump are generally provided for each electric plants. Four-cycle engines tend to be heavier turbine exhaust condenser. These pumps are normally and costlier but more efficient than two-cycle engines. provided by the shipbuilder but may be furnished with Most marine engines above 1000 hp are of the two-cycle the turbine generator. Pipe connections may be made type. Two-cycle engines are generally equipped with so that during emergencies the generator turbine exhaust attaclyxl positive-displacement blowers to supply scavcan flow to the main propulsion turbine condenser. The enging air to expel gases from the cylinder a t the end of generator turbine condenser may be arranged to receive the exhaust stroke. I n addition, both two- and four-

612

MARINE ENGINEERING

cycle engines may be equipped with turbochargers (driven by gears, exhaust-gas turbine,, or a combination of both) to increase the engine capacity and provide improved fuel economy. Most marine engines of 2000 hp and over are turbocharged to reduce their size and weight. It is important that provisions be made for sufficient combustion air, either by running air ducts from the outside directly to the engine air intake, or by means of the room ventilation system. Additional room ventilating air also must be provided for removing engine waste heat not removed by the lubricating oil and jacket water systems. The ductwork must be checked for pressure drop from the outside atmosphere to the engine; usually not over 6 in. of water is allowable to obtain guaranteed engine performance. For the same reason the exhaust-gas pipeline from the muffler to the atmosphere should be sized to obtain a back pressure at the muffler outlet not exceeding about 16 in. of water. The complete marine engine includes additional attached and unattached equipment as follows: 1 Fuel control system consisting of an engine-driven fuel pump, duplex filter, suction strainer, and injector control lever for manual starting, stopping, and emergency speed control. 2 Lubricating oil system consisting of an engine driven oil pump, full-flow filter with by-pass relief valve, strainer with relief valve, and cooler with by-pass relief valve. When dry-sump engines are used, the lubricating system includes a scavenging pump with suction strainer and by-pass relief valve. 3 Piston cooling oil pump. 4 Freshwater systems for diesel-generator sets larger than 350 kw consisting of an expansion tank, enginedriven water pump (some engines may have two pumps) cooler, and automatic water temperature regulator. Smaller sets generally have a radiator and fan for cooling fresh water. 5 Exhaust system consisting of a dry, spark-arresting type muffler and watercooled or insulated exhaust header. For 350-kw generator sets and larger, the engine is usually equipped with an exhaust temperature indicating system consisting of a a t of thermocouples, a selector switch, and galvanometer. One thermocouple is installed in the exhaust of each cylinder. The selector switch and galvanometer may be mounted on the gageboard. 6 Starting air system consisting of a starting motor (two motors may be necessary to accelerate large engines) or air distributor for sequential admission of air to the engine cylinders, strainer, and air control valve. A solenoid-operated valve may be provided for remote engine starting from the switchboard. Starting air systems are generally designed to operate a t pressures of 125 to 250 psi. Starting air tanks and the air compressors are normally furnished by the shipbuilder. Diesel generators rated at 500 kw or less generally use either an electric or hydraulic starting system. Capacity for a t least ten successive starts, beginning with a cold engine, is provided with either method. Either a centrifugal or

pressure-operated device is provided to prevent inadvertent attempts to initiate cranking after the engine has started. These devices or a solenoid-perated valve are used to automatically stop cranking after the engine has started. Also, means are provided to prevent unintended starting during maintenance of the set; for electrically started engines, a disconnect switch is provided in the starting clcuit; for hydraulically and pneumatically started engines, a cutoff valve is provided in the associated piping system. 7 Air intake filter-silencer. 8 Oil relay type constant-speed governor or electrohydraulic load-sensing speed governor when small speed variations and quick response are desired. 9 Speed adjusting (synchronizing) device designed for local manual adjustment and an electric motor or potentiometer for remote adjustment from the switchboard when synchronizing generators. 10 Overspeed trip device which upon overspeed of not over 15 percent will close the fuel racks or combustion air supply to shut down the diesel. 11 For ship service diesel generators, an interlock switch for actuating the generator circuit-breaker tripping device to disconnect the generator from the switchboard when the engine is shut down due to overspeed. 12 Gageboard including, as a minimum,. pressure gages for the freshwater and seawater pump discharges, fuel oil and lubricating oil flter inlets and outlets, lubricating oil strainer inlet and outlet, scavenging air, starting air, and also thermometers for the fresh water and lubricating oil from the engine. 13 Manual engine-turning gear for turning small engines or an air-motor type turning gear, with an interlock, for turning large engines. 14 I n addition to the aforementioned equipment, turbocharged engines are equipped with a scavenging turbocharger and turbocharger air intake cooler. Tubetype coolers are used for cooling the lubricating oil and fresh water for diesel generators larger than 350 kw. Seawater is used as the cooling medium in the freshwater cooler and fresh water is used as the cooling medium in the lubricating oil cooler. For sets rated 350 kw or less, fresh water is generally cooled by a radiator and engine-driven fan. If diesel generator sets are orderedior paralleling with existing generators or with new sets procured from another manufacturer, the required speed regulations and electrical characteristics must be specified. c. Gas Turbine Prime Movers. Gas turbine driven generator sets are available for ship service or emergency power and a limited number have been installed. Gas turbines are smaller and Lighter than comparable steam: turbines or diesel engines. Since gas turbines operate a t very high speeds, they are usually noisier than steam turbines. The starting time for gas turbines is normally 30 to 40 sec, but it can be reduced below 10 sec when used for driving emergency generators. Batteries or compressed air may be used for starting.

I

ELECTRIC PLANTS

d. Electrical Characteristics. Except for small loads supplied from batteries, all direct-current loads are supplied by a-c/d-c motor generator sets or by rectifiers. Rotating field type a-c generators are used for ship service power. The generators are rated at 450 volts, 3 phase, 60 cycles. Stator windings may be either delta- or wye-connected but usually are the latter; only three main terminals are required per generator. The inherent voltage regulation of a-c generators is comparatively wide due to the high synchronous reactance of the windings. This is an advantage inasmuch as the synchronous reactance limits short-circuit current, but in order t o maintain the required voltage regulation an automatic voltage regulator must be used with each machine and the degree of regulation depends on the sensitivity of the regulator. Generators may be dripproof protected or totally enclosed. If totally enclosed, the generator will be equipped with a double-tube air cooler using seawater as a cooling medium. Silicone insulation should not be used for totally enclosed generators unless the slip rings are located outside the generator enclosure; this is to prevent abnormal brush wear and increased slip ring maintenance. If the design and arrangement of the generator are such that circulating currents may be expected in the rotor shaft, means (such as the use of insulated bearings) should be provided to prevent circulating currents from passing between the journals and the bearings, as the babbitted bearing surfaces may otherwise be destroyed. Ship senrice generators that weigh more than 1000 lb, excluding the shaft, and all emergency generators should be provided with electric space heaters to prevent moisture condensation during shutdown. Generators rated a t 500 kva and above should be provided with resistance-type temperature detectors embedded in the stator windings, The temperature indicating instrument should be located conveniently, .preferably on the switchboard generator control panel. An excitation system is procured with each a-c generator. Two types of rotating exciters are the d-c exciter and the a-c brushless excitkr. Both of these exciters are coupled to the generator shaft and are similar in outward appearance. In place of- the commutator on the d-c exciter, the a-c brushless exciter has a solid-state three-phase rectifier mounted on the generator shaft to provide the d-c excitation for the generator field. The a-c brushless exciter responds faster than the d-c exciter. A third type of exciter in general use is the static excitation system. This system eliminates the necessity of rotating components and has a faster response than either type of rotating exciter. ,Selection of the exciter and voltage regulator should be coordinated to obtain the desired recovery time. The a-c. brushless exciter and the static excitation system have in general replaced the rotary amplifier type of exciter. e. Emergency Generators. Emergency generators provide a power source independent of any other equip-

61 3

ment on the vessel and are usually diesel driven. The requirements outlined in the preceding paragraphs for diesel engine prime movers apply equally to emergency sets, except as noted in the following. Each engine is equipped with a self-contained cooling system requiring a radiator and fan. Tf ventilation ducts are installed to and from the engine radiitor, the radiator fan must develop enough head to force the cooling air through the ducts. Motor7:;dperatedlouvers are generally installed in supply and exhaust duct terminals which are exposed to the weather. The vent motors are energized from the generator side of the generator circuit breaker to insure that the louvers are open when the diesel engine is in operation. Emergency generator units should be arranged to shut down automatically upon loss of lubricating oil pressure, dangerous overspeeding, or release of carbon dioxide in the emergency generator room. An audible alarm device should be provided that will sound in the event of low oil pressure or a high cooling-water temperature. Engines are generally arranged to start automatically upon failure of ship service power. If battery-started, a voltage-sensitive relay with contacts that close when the ship service power fails is used to energize the control circuit of the starting motor; if hydraulically started, a loss of ship service power will de-energize a solenoidoperated valve to initiate the starting process. In either case, means must be provided for automatically rendering the starting devices inoperative after the engine has attained firing speed. Devices for manual starting control are required with each type of starting equipment for test purposes. Emergency generators are not required to operate in parallel with the ship service generators, except where closed transition transfer with proper synchronization with ship service generators is desired (to prevent power outage when transferring power sources). An automatic voltage regulator should be provided with each emergency generator. The emergency generator should have the same voltage rating as the ship service generators and should be of sufficient capacity to permit "cold" starting of the ship's main power plant. f. Special Generator Arrangements. A variable generator frequency is obtainable by using special control equipment. Variable-frequency generators may be desirable where speed control of several large motors is required. Generally, it is more economical to use d-c motors supplied by an a-c/d-c motor generator set for most variable-speed shipboard applications. Multiple unit generator sets are available, eack consisting of two generators driven by one steam turbine. Diesel generator sets consisting of one generator driven by two diesel engines have been installed on some vessels. In case one engine fails, it may be declutched and the generator can continue to deliver approximately one half of its rated output. g. Voltage Regulation. A direct-acting type of voltage regulator employing a mechanical regulator element may be used for controlling the field of d-c

614

MARINE ENGINEERING

rotating exciters or very small auxiliary generators. The regulator element may be a torque motor, solenoid, or electro-dynamometer energized by the generator voltage and restrained in motion by a spring. Rheostatic elements are automatically operated field rheostats of which there are two types. One consists of a motoroperated face plate rheostat controlled by means of contacts on the regulator element and the other consists of a resistor with numerous steps cut in and out by mechanical linkage with the regulator element. Both types have antihunting or damping devices to prevent fluttering with small changes in load and overregulation with large sudden changes in load. Either of two types of static voltage regulators may be used depending on the type of excitation system. When a static excitation system is used, the voltage regulator senses the generator output voltage which is rectified and applied to the control winding of a magnetic amplifier. The output of the amplifiers is impressed across the control winding of three saturablecurrent potential transformers, one per phase. The outputs of the transformer secondaries are rectified and impressed across the generator field. If an &c brushless exciter is used, the exciter field current is supplied by a static voltagr: regulator which senses generator output voltage. An error voltage is impressed across a reactor which becomes saturated, conducts, and fires a silicon-controlled rectifier to provide current to the exciter field. The average exciter field current is determined by the point a t which the rectifier fires during each positive half-cycle. Provisions are made in static regulators to provide sufficient excitation that will result in fault currents large enough to ensure selective tripping of overcurrent devices during short-circuit conditions when the generator voltage is zero. The regulators described in the foregoing are suitable for use with generators operating in parallel, each generator requiring an individual regulator. For parallel operation, a compensating effect to reduce wattless

ELECTRIC PLANTS

current is obtained by making use of the potential generated in the secondary of a current transformer, the primary of which is in series with one of the main leads. Each regulator, except for some emergency generator applications, ingludes a transfer switch for cutting out automatic operation and provides for manual control of the generator voltage by means of a rheostat. h. Generator Terminals. The preferred run of cables between the generator and switchboard is downward, under the generator platform, and up behind the switchboard. The preferred location of the terminals is therefore at the bottom of the generator, unless the generator is of such small size that connections may he made inside a standard terminal box mounted on the side of the generator frame. All generator terminals should be protected against accidental contact and mechanical damage. If terminals are located on the top or side of the generator frame, they should be protected by a watertight enclosure with removable covers furnished by the manufacturer. Where cables enter a terminal enclosure on the top or side of the generator frame, the enclosure should be provided with terminal tubes. For terminals located on the bottom of the generator frame, the necessary protection, usually expanded metal, is furnished by the shipbuilder. Additional terminals are required for exciter and imbedded temperature detectors and generator space heaters, if used. Any such miscellaneous terminals should be located and protected in a manner similar to the main terminals. All generator power connections should be silver-plated. The number, type, and size of terminal lugs to accommodate the conductors should be specified and should be clearly marked by the generator manufacturer for identification. Particular care should be exercised when terminals are brought out and marked to assure that the phase rotation with respect to the terminal marking will be the same on all generators, including emergency generators.

Section 3 Switchboards and Panels %

The following does not cover in detail the design and construction of switchboards since these details are the responsibility of the switchboard manufacturer. The procurement of marine switchboards does, however, require that the shipbuilder preparefor the guidance of the switchboard manufacturer-detailed descriptive specifications giving complete information on the required type and arrangement of the switchboard, the number, rating, and wire size of feeder circuits including bus ties, and any limits to the height, length or depth of the switchboard imposed by its location in the vessel. An elementary diagram of the electrical distribution system should also be given. 3.1

General.

The first step in the preparation of switchboard descriptive specifications is a thorough study of the ship's requirements. The ship's requirements determine the number, rating, and type of generators to be controlled and usually impose specific requirements regarding the switchboards. A review of the requirements of all applicable rules is also advisable. 3.2 Special Requirements. The following features should be considered in connection with switchboards for installation on shipboard. Switchboards should be located in dry areas and should be accessible from the front and rear. Also, switchboards should be located as far inboard from the ship's sides as practicable. The

usually restricted space on shipboard requires careful study of the switchboard assembly drawings, hull structural drawings, and machinery arrangement plans to ensure space for installation without interference frdm girders, beams, stanchions, bulkhead stiffeners, and major equipment. The space in front and rear of switchboards should be unobstructed and adequate for operation and mainte nance; minimum clearances are 36 in. in front and 30 to 36 in. a t the rear (see regulatory rules for specific requirements). Since the space behind switchboards should not be accessible to unauthorized personnel, an enclosure is usually provided which extends from the ends of the switchboard to the ship's structure with doors arranged for locking. When this arrangement is not feasible, protective enclosures are mounted on the switchboard rear framework. Switchboards should be located as close to their associated generators as is practicable so as to keep the length of cables to a minimum. Space permitting, neither steam, water, nor oil lines should be located over or dose to the switchboards. Under all conditions, pipe joints should be located away from switchboards. Drip shields should be provided a t the top of all switchboards to protect against dripping from pipes and also against falling objects. Ventilation ducts should not discharge air directly on the suritchboard. Insulating mats or gratings should be provided on the deck in front and rear of switchboards to insulate personnel from ground and to prevent slipping. These mats or gratings should extend the entire length of the switchboard and be of sufficient width to suit the operating space. Nonconducting handrails attached to the front of switchboards are provided; usually these handrails are horizontal. Nonconducting guardrails are also provided a t the rear of switchboards. The switchboard and its component parts as finally installed must be capable of withstanding shipboard vibration without damage or faulty operation and should operate successfully when inclined a t an angle of 30 deg in any direction from the vertical. Special consideration should be given to adequately supportingtthe bus bars. If required, switchboards may be braced to a bulkhead or the deck over. However, overhead bracing should be flexible to allow deflection of decks without causing the switchboard structure to buckle. 3.3 Rating and Characteristics. The rating and characteristics of the ship service generator switchboards and emergency generator switchboards are usually fixed by the rating, type, and arrangement of the electric plant. Three-phase, 450-volt, 60-cps generation and distribution systems are standard for both passenger and nonpassenger vessels. The quantity and kilowatt rating of the generators are dependent upon the total connected loads and demands of the electrical system. Direcb

current power supplies 'are often provided for cargo winches, and other loads requiring precise speed control. This power source will normally consist of motor generator sets with individual control and distribution facilities separate and apart from the main switchboards. Switchboards, both' maia and emergency, usually contain a 3-phase, 120-volt, 60-cps distribution section, energized via transformers, having suitable kva ratings, from the local en-volt, Bphase, 60-cps bus. 3.4 Types bf Switchboards. 'There are two general, types of switchboards, dead-front and live-front. The applicable rules require that dead-front type switchboards be used for all a-c applications where the voltage to ground or between poles is in excess of 55 volts and for all d-c switchboards where the voltage to ground or between poles is in excess of 250 volts. A dead-front design provides for all energized parts to be enclosed within the switchboard structure, whereas a livefront design permits surfacemounted fused lever switches, circuit breakers, and instruments. A deadfront design offers advantages involving personnel safety and equipment protection, and is usually a specific requirement of ship specifications for all marine switchboards. On dead-front switchboards, open-frame circuit breakers, rheostats, and other heavy equipment requiring front-of-board o~eration are mounted on suitable supports fasteneb to interior angles with only the operating handles or knobs projecting through the front hinged panel. Distribution circuit breakers are mounted on suitable support plates fastened to interior angles with only the operating handles projecting through the front panel. Instruments, indicating lights, and control switches are semi-flush mounted on hinged front panels. 3.5 Lighting of Switchboards. Switchboard illumination may be provided by the compartment lighting system. However, a preferred arrangement is the use of lighting fixtures mounted on and under the overhang of the switchboard drip shield. The lights, of which there may be several, are connected in part to the emergency supply so as to provide adequate illumination upon loss of normal supply. 3.6 Arrangement of Switchboards. The usual practice is to provide a panel for the control of each generator and additional panels in number as required for circuit breakers and switches controlling power distribution. With switchboards of small and medium size, the generator panels may be a t one end of the switchboard and all feeder panels to the right or left. For large switchboards, a considerabId9avings in'cost of the main bus may be effected by locating the generator panels in the center with feeder panels of approximately equal working load on each side. With this arrangement no part of the main bus carries more than approximately one half of the total working load. 3.7 Switchboard Applications. Switchboards are normally used in the following shipboard applications: a. Main Generator and Distribution Switchboard. The main switchboard provides for the control, protec-

ELECTR1IC PLANTS

61 6

tion, and paralleling of the local ship service generators and control and protection of the ship service power distribution system. Usually only one main switchboard is required on all but large passenger vessels and special-purpose vessels. For large and special applications two or more main switchboards, connected together through bus ties, may be installed depending on the number and location of main generators. ~ e a n sof connecting the main ship service bus to shore power and also to the emergency switchboard are provided through a shore-power circuit breaker and a bus tie circuit breaker respectively, each located on the main switchboard. Normally main switchboards consist of a generator panel for each connected ship service generator, a shore power and bus tie panel, and distribution panels in size and number as required. Figure 3 is a typical main switchboard one-line diagram. Figure 4 is a typical main switchboard generator control panel one-line diagram. Each generator control panel should include, but not be restricted to, the following : 1 A trip-free circuit breaker with separate overcurrent trips for each pole except that trips are not required for the neutral of dual-voltage systems. The circuit breaker should be arranged to open all three poles with a predetermined load existing on any one or combination of poles. When two generators are installed, each generator breaker should have inverse-time overcurrent trips; for three or more generators arranged for parallel operation, each generator breaker should have inversetime and instantaneous overcurrent trips. (Inverse time tripping means that the opening time of overcurrent devices decreases as the magnitude of the current increases.) For large generators the circuit breaker should be electrically operated to provide personnel protection and for quick closing during paralleling. 2 An unfused disconnect device which will completely disconnect the generator and its circuit breaker from the bus. This device is not required when the generator circuit breaker is of the drawout type. 3 An ammeter, with a selector switch to read the current of each phase. 4 A voltmeter, with a selector switch to read each phase of the generator and one phase of the bus. Also, one of the generator voltmeter switches should provide for reading each ~ h a s eof the shore connection. 5 A sGchro&ope and synchronizing lamps with a selector switch to provide for paralleling ship service generators including the emergency generator when closed transition transfer between the main and emereencv " switchboards is desired. 6 A control for prime mover speed for paralleling generators and frequency adjustments. (The speed control may be located on the engine for emergency switchboards.) 7 An indicating wattmeter. 8 A frequency meter with a selector switch to connect

-

-

:

MARINE ENGINEERING BUS

to any generator. This may be common to all generators. (This is required for the local generator only on the emergency switchboard.) 9 Space and mounting for generator field and exciter rheostats. 10 A switch and indicator light for the generator space heater supply. 11 A double-pole field switch with discharge clips and resistor. For generators with variable-voltage exciters or rotary-amplifier exciters, each controlled by a voltage regulator unit acting on the exciter field, and for generators equipped with static excitation and regulating systems, the field switch, the discharge resistor, and the generator field rheostat may be omitted. 12 A voltage regulator complete with accessories. 13 A switch for cutting out the voltage regulator. 14 A white light to indicate that the generator is running. 15 Adequate means for ground detection. 16 The necessary current and potential transformers. 17 Current and potential test receptacles for use with portable instruments (only on ship service switchboards). 18 One temperature indicator and selector switch, when the generators are rated a t 500 kva and above. 19 A reverse-power relay for each generator, when two or more generators are to operate in parallel. b. Emergency Switchboards. The emergency switchboard provides for the control and protection of the emergency generator and the emergency power, lighting, and interior communication systems. Usually only one emergency switchboard is installed; however, for safety reasons large passenger vessels may have two emergency sources of power, thus requiring two emergency switchboards. Emergency switchboards normally consist of a generator control panel, a bus tie and distribution panel and 24-volt, 120-volt, and 450-volt distribution panels in size and number as required. Emergency generator switchboards include those devices noted in the foregoing for main switchboard generator control panels plus the following: 1 A white light to indicate that the normal supply is available. 2 A green light to indicate that all devices are "set up" for automatic operatio& (This indication should be extended to the ship service switchboard to alert the operator when the emergency generator controls are not properly positioned for automatic startup.) 3 A normal supply circuit breaker, unless the automatic bus transfer is of the contactor type. 4 A feedback switch when the automatic bus transfer is of the contactor type. Except under emergency conditions, the emergency switchboard serves as an extension of the ship service power distribution system with its main bus supplied from the ship service main switchboard through a bus tie and an automatic bus transfer device. This bus transfer device usually consists of two circuit breakers, the emergency generator breaker, and the bus tie breaker,

i

I

I

t TOSHORE PWR TERMINALS

I

t

TOEMER SWITCHBOARD

LEGEND NO. I CIRCUIT BREAKER -- GEN. GEN. NO. 2 CIRCUIT BREAKER C - SHORE PWR CIRCUIT BREAKER POTENTIAL TRANSFORMER. .CT. - CURRENT TRANSFORMER

0 -NORMAL SUPPLY CIRCUIT BREAKER E -0ISTRIBUTION CIRCUIT BREAKER F -LOCAL 4 5 0 / 1 2 0 VOLT XFMR ClRCUlT BREAKER FU FUSE

A

-

PT-

~

Fig. 3 Typical main switchboard one-line diagram

r 450 VOLT, 3 PH, 60 CPS

GENERATOR BUS

I

1

SYN LTS

Q

Q

1

T

C

TO 120 VOLT BUS

SHUNT TRIP

'

------LEGEND VR -VOLTAGE REGULATOR FM -FREQUENCY METER AM -AMMETER SYN SYNCHROSCOPE PT -POTENTIAL TRANSFORMER WM - WATTMETER TM - TEMPERATURE METER CT -CURRENT TRANSFORMER VM -VOLTMETER GOV -GOVERNOR B GENERATOR CKT BKR A -GENERATORCKTBKR FU -FUSE . Fig. 4 Typical main switchboard generotor control panel --line diagram

-

-.

MARINE ENGINEERING 450 VOLT, 3 PH, 6 0 CPS EMERGENCY BUS

619

ELECTRIC PLANTS

I20 VOLT, 3 P H 6 0 CPS. BUS

24 VOLT DC BUS

I

?

?

-

4 5 0 VOLT, 3PH, 6OCPS / F!NAL EMER BUS .-

T

I T

?A

?

TWO BREAKER TYPE ABT INTLK

I

4 5 0 1 120 VOLT TRANSFORMER

t

4 5 0 VOLT POWER

120 VOLT, 3 WlRE FINAL EMER BUS ? ? T

AC/DC MG OR CONVERTER CONTINUOUSLY OPERtTlNG

'

I

450/120 VOLT TRANSFORMER

FINAL EMER LOADS

120 VOLT, 3 WlRE TEMP EMER BUS

1

I

b t AUX

TO

GEN HTR

I

EM GEN CKT BKR AUX CONTACTS

I I

9-' I

EMER LTG

?VITAL AC LOADS

I I

I

I

LEGEND A

- QEN CIRCUI,T BREAKER

B -NORMAL SUPPLY CIRCUIT BKR C T -CURRENT TRANSFORMER WM- WATTMETER U V UNDERVOLTAGE TRIP VR + VOLTAGE REGULATOR

-

-LTG XFMR CIRCUIT BKR C D - 4 5 0 VOLT DISTR CKT BKR E -120 VOLT DlSTR CKT BKR F U -FUSE FM -FREQUENCY METER HTR HEATER

-

-

F BATT CHARGER CKT BKR G 2 4 VOLT DC DISTR CKT BKR PT -POTENTIAL TRANSFORMER AM- AMMETER VM- VOLTMETER SW- SWITCH

CONTACTOR TYPE ABT

LEGEND AUX SW8 ON MN GEN BKRS

AM -AMMETER VM -VOLTMETER F M -FREQUENCY METER WM -WATTMETER MG -MOTOR GENERATOR P T -POTENTIAL TRANSFORMER C T -CURRENT TRANSFORMER ABT- AUTOMATIC BUS TRANSFER

/

i NORMAL SUPPLY Rg. 5

Typical cargo ship or tanker emergmcy switchboard me-line diagram

SKETCH "A"

-

FU FUSE A GENERATOR CKT BKR B NORMAL SUPPLY 450/120 VOLT XFMR CKT BKR C D -DISTRIBUTION CKT BKRS (AS REQ'DI E -MOTOR GENERATOR CKT BKR , F BATTERY CHARGER CKT BKR

-

-

Rg. 6

- AUTOMATICALLY TRANSFERS EMERGENCY LIGHTING TO BATTERY UPON LOSS OF NORMAL SUPPLY @ - AUTOLATICALLY TRANSFERS VITAL AC 6 0 CPS LOADS

ABT @

ABT TO MG UPON LOSS OF NORMAL SUPPLY ABT @ -AUTOMATICALLY TRANSFERS VITAL DC LOADS TO BATTERY UPON LOSS OF NORMAL SUPPLY (MG SET)

Typical passenger vessel emergency switchboard me-line diagram

4

both of which are electrically interlocked so that only one breaker can be in the closed position at a time. The bus transfer device is arranged to initiate an automatic start of the emergency generator set upon failure of the normal power supply. When the emergency generator ig up to rated voltage, the bus transfer operates so as to transfer the emergency bus from the normal supply to the emergency generator. Upon restoration of the normal power supply, the emergency loads may be manually transferred to the normal supply and the emergency generator manually stopped. Retransfer to the normal supply, and particularly shutdown of the emergency generator, should be done only after the continued operation of the normal supply is reasonably assured. See Figs. 5 and 6 for typical emergency switchboard one-line diagrams. On some small-capacity installations, the bus transfer device is a contactor type in lieu of a circuit-breaker type principally because of cost considerations. Functionally the operation of a contactor type is the same as the conventional two-breaker type transfer; however, the arrangement is different as noted in Fig. 5.

When the normal supply is not available, it is often desirable to feed emergency power back into the ship service system. To satisfy this condition when using a circuit-breaker type transfer, provision is made to defeat the electrical interlocking feature, thus allowing both circuit breakers to be in the closed position a t the same time. When using contactor-type transfers, a manually operated feedback circuit breaker is provided to complete r the normal a circuit between the emergency g e n e ~ t oand supply bus tie. When so arranged, electrical interlocking should be provided to prevent inadvertent paralleling of the emergency and ship service supplies. For a passenger vessel the emergency switchboard has two sets of buses, one designated 'final emergency" and one designated "temporary emergency. " The "final emergency" buses (450 and 120 volt) are normally energized through a bus tie from the main switchboard as described previously. The "temporary emergency" buses (120-volt Bwire, 120-volt 1-phase, and 120-volt d-c) are normally supplied from the "final emergency" buses through associated transfer switches and are automatically transferred to a storage battery supply

upon failure of the normal power. This provides instantaneous restoration of the emergency power supply, either directly from the battery or by means of d-c/&c conversion equipment supplied from the battery, to those loads requiring an uninterrupted power source. See Fig. 6 for a typical emergency switchboard one-line diagram for passenger vessels. c. Load-Center Switchboards. Load-center switchboards are essentially remotely loc&ted sections of the main switchboard distribution section. They are supplied from the main switchboard via a bus feeder and in turn supply power to local lighting and power loads. Load centers are centrally located regarding the loads supplied for reasons of convenience and economy and are normally installed on only large passenger vessels having considerable power requirements located throughout the vesael. (Figure 18 is a one-line diagram of a typical load-center application.) 3.8 Selectivity. Selectivity provides for maximum continuity of service under fault conditions through the selective operation of various protective devices, that is, the isolation of a fault with the least interruption of vital senrices.

Selectivity is obtained through the coordination of trip devices of the various breakers in the system. The time bands, both long and short time, of the open-frame type of breakers require special attention for coordin& tion, especially when the spread between breaker trip ratings is very narrow. When fused breakers are employed, very careful attention must be given to the coordination of fuses with the breakers, and with other breakers in, the system. Complete information with regard to generator maximum and sustained fault current and system impedance must be available in order to develop a properly coordinated system. For a more detailed discussion of selective systems see Sub,%section 6.7. 3.9 Circuit Protective Devices. Protective -devices such as oircuit breakers, fuses, reverse-power relays, and current-sensitive relays are installed on switchboards to provide protection against faults in the electrical distribution system. These devices are applied so as to isolate any fault with the least possible portion of the system being interrupted; the arrangement should be such that the generator circuit breaker is the last to open under fault conditions. Each protective device (circuit

620

MARINE ENGINEERING

breaker) must have an interrupting rating adequate to safely interrupt the maximum fault current obtainable at the point of application. Generator and bus tie breakers and breakers feeding combined loads in a selective system are of the openframe type and have long-time and short-time trips. Instantaneous trips may be used in those cases where the circuit breakers have a short-time rating below the interrupting rating, provided they do not defeat selectivity. Breakers feeding individual loads in the system should be of the molded-case or open-frame type with the only requirement being that the continuous and interrupting ratings are adequate for the application. When three or more generators are to operate in parallel, the generator circuit breakers should have instantaneous trips which are set a t a .value in excess of the maximum fault current obtainable from an individual generator. Fuse selection is based on system characteristics (voltage and current) and speed response (standard or time-delay) required a t the point of application. All fuses should be of the nonrenewable cartridge type and capable of interrupting the available fault current. Reversepower relays are provided to prevent a generator from operating as a motor when paralleled with another generator. This relay trips its associated generator breaker when power flows from the line to the generator in lieu of from the generator to the line. Usually the relay is set t o initiate generator breaker tripping within 10 sec when reverse power is approximately 5 percent of the generator rating. I t is sometimes necessary to employ current-sensitive relays for tripping circuit breakers ,at a predetermined current. This becomes of special importance when arranging for selectivity between the emergen:y generator gnd main switchboard bus tie breaker during feedback operation. Under normal conditions (with the emergency switchboard energized from the main switchboard), the bus tie breaker is coordinated with the main generator breaker, and has a relatively high trip value. During feedback operation it is necessary to trip the bus tie breaker should the combined load (both main and emergency) on the emergency generator approach a value likely to exceed the emergency generator rating. This is required so as to provide for continuity of emergency supply to emergency loads. The tripping scheme would normally consist of current transformers and current-sensitive relays properly coordinated and arranged to monitor the emergency generator total current and act on a shunt trip of the main-emergency bus tie breaker. The trip circuit would be electrically interlocked so as not to be effective a t any time other than when operating under a feedback condition. To avoid interrupting vital circuits as a result of overload tripping the generator circuit breakers, nonvital loads may be arranged for automatic tripping when the total load on any generator or bus tie circuit exceeds a predetermined value. Usually, the loads to be tripped

are connected to a common bus or buses; tripping the breaker feeding the bus thus disconnects all loads commonly connected. In lieu of the foregoing, nonvital loads may be sequentially tripped via a multiple-contact timing relay. The contacts are arranged to close serially a t predetermined intervals, and the closing of each contact trips one or more circuit breakers serving nonvital loads until the overload is reduced to an acceptable value. With the overload cleared, the relay contacts open and the timing relay resets. The circuit breakers to be tripped must have either a shunt or undervoltage trip device, undervoltage being preferred because of its fail-safe characteristic. Current transformers of the proper rating are required for each generator or bus tie circuit to be monitored, and a current-sensitive relay is required for each breaker or group of breakers to be tripped. The relay may be instantaneous or timedelayed so as to not initiate tripping on momentary overloads. 3.10 Types of Panels. The panels most commonly installed aboard ship for the many specialized service requirements are of the following types: a. Distribution Panels. Lighting and power system distribution panels have the same'function as do loadcenter switchboards, i.e., supply power to local lighting and power loads respectively. They are supplied from either the main switchboard, emergency switchboard, or a load-center switchboard. Distribution panels are normally of a dripproof construction and located in dry areas central to the loads they supply. Distribution panels should not be accessible to unauthorized persons. They are surface mounted except in passenger, crew, and public areas in which case they are flush-mounted in way of joiner work. Panels are fitted with multipole switches or circuit breakers having a pole and an overcurrent protective device for each associated circuit conductor. In general, 440-volt a-c panels are restricted to a maximum of 12 three-phase circuits, and 115-volt lighting panels are restricted to a maximum of 14 threephase circuits (42 overcurrent devices). Lighting papels are normally arranged for a threephase supply and single-phase distribution; power panels are normglly arranged for a three-phase supply and three-phase distfibution. b. Alarm Panels. Alarm panels are commonly provided for monitoring various criticaT checkpoints associated with the ship service genepators, emergency generator, generator prime movers, and vital propulsion plant auxiliaries and systems. Checkpoints may be monitored a t one common panel or separate panelstmay be provided for specific systems. Alarm panels should incorporate audible and visible indication of a system failure or "off-normal7' condition. Provisions should be made for silencing the audible alarm with the visual indicator remaining "on" to indicate a standing fault. Common practice is to have a "flashing" light indication for the initial fault and a "steady" light

ELECTRIC PLANTS

I

I

I

.

I I

for indication of a standing fault. Suitable means should be provided for indicating an "out of service" condition and for simulation of a fault for testing each system. Alarm circuits should be designed for operation from a 24-volt or a 120-volt supply unless the circuits are simple, in which case a 6- or 12-volt supply should be satisfactory. Each alarm panel should have a control power transformer, power supply fuses, power-available light, and power supply control switch. In some instances remote power supplies are provided for specific systems, and, when so arranged, power-available indicating lights should be provided for each such system. c. Test Panels. A panel arranged to provide test voltages at all values utilized on the vessel should be installed in a convenient work area for the purpose of testing electrical appliances and components. Normally, the *C power supplies (440 volt and 115 volt) to the test panel are from the ship service systems and the d-c supply is from a rectifier of the proper voltage and rating. Complete instructions for operating all test devices should be mounted either on or adjacent to the test panel. d. Special Consoles. A console arrangement provides for the grouping of selected navigation, communication, and/or propulsion plant indicators and controls on

or within a common enclosure usually having a sloping desklike top. Typical examples are wheelhouse command consoles and engineers' operating station consoles provided on vessels having automated propulsion plants. *Typicalexamples of these types of consoles are given in Chapter 21. On vessels having centralized control for propulsion plants, it is desirable that' the h i p service generator control be adjacent to or in the vicinity of the propulsion plant controls. Unle~sthe eritire ship service switchboard can be conve+6ntly located, the generator controls are grouped on a control unit apart from the switchboard so as to permit a choice of location. The control unit may be a vertical type of the same design as if it were part of the switchboard, or it may be of the console type. A generator console type of control unit permits grouping the generator circuit breaker control switches and instrument control switches on the console sloping top. Instruments, indicator lights, and less frequently used control switches are mounted on the vertical area of the console. Rheostats and voltage regulator equipment are usually located in the bottom section of the console. A console type of design is preferred when several generators are to be controlled from one location or when a large number of controls are required for remote devices.

Section 4 Power Eq~~ipment 4.1 Genered Requirements for Motors and Controls. types of enclosures and methods of ventilating motors General recommendations for the construction and most commonly used are as follows: application of motors and control apparatus for marine 1 Open, self-ventilated-used only where an adequate service are contained in the IEEE Standard No. 45. enclosure is provided by the housing of the driven Certain specific requirements are also contained in the machine. This type of enclosure should have ventilaregulations of the classification societies and in the tion openings which permit passage of external cooling United States Coast Guard Code of Federal Regulations air over and around the motor windings. -Title 4Mhipping, Parts 1 to 149. 2 Dripproof protected, self-ventilated-used for most Motor and control equipment for marine service may applications in dry, sheltered locations. This enclosure be classified into several standard types relative to is so constructed that drops of liquid or solid particles mechanical and electrical characteristics. The proper falling on the motor at any angle not greater than 15 application of motors and controls for shipboard deg from the vertical cannot enter the motor. The auxiliaries involves, therefore, the selection of those ventilating opefLings are riormally protected with wire standard types having mechanical and electrical charac- screen, expanded metal, or perforated covers to prevent teristics that are most suitable for the location and na- personnel from contacting electrical parts. These ture of the driven auxiliary. covers also keep out rats and mice that might use the The ratings and characteristics of electrical motors and equipment as nesting places. controls fop a typical singlbscrew cargo liner are listed 3 Totally enclosed, fawcool&-generaly used for in Table 2. motors in spaces where lubricating or fuel oils are present; they are also used for applications subject to splashing, 4.2 Mechanical Characteristics of Motors spraying, or hosedown. This type of enclosure prevents a. Enclosure and Method of Ventilation. A wide the free exchange of air between the inside and outside variety of enclosures and methods of ventilation is' of the housing but is not sufficiently enclosed to be available for motors in marine service. The specific termed airtight. types selected depend on the particular environmental 4 Waterproof, nonventilated-used for practically all condition to which the motor will be subjected. The motors mounted on weather decks, or where heavy

I

I

M A R Y E ENGINEERING

622

ELECTRIC PLANTS

1

L+s

asga 3

L

5

.e

4

a 0

g

0)

L

"

I

r B $

, "

2

FX'B $3 'E: 13 e'&;'z 3% 2e 1 0 0 0 0 0

aaa aa .pba .+a

a a .a

aaaaaa .4a.4a.4a

-48

mmaaa a aa 4aaa4a

YYY ?y 228 st

Y =??V) 1"

a

ge

rn-m-ga m

U

a

3.p Bm

U

B a

U

8 &

UUUU

Ha Ba Ha Ha

a .a .a .a .a.+a.a

UUCU

UU

U

U

U

H l H B 18 B B H aaaa aa a a a

aaaa aa

a

a

a

B

a

I -5

a aaaz 2aaa a,: a a a a

-

?* qr w55 =?r 2 w Rw-*

o a ,?S urw> a-

d

e4

\

,A

j l

? ? w

d

i3g

U

U

B B a a

5

4

623

9599 5 .9 9 5% 5% 59

-.. . " f 9

-$.$

1 xa 1a a' ~a :a1a1 12a a' Ea: a~ ~a . aB ' E1a : 2a a2x1 a

Pi Pin PI

g:ga3

1 2 2 '81 a a

a aa a .a.a

a ~ 28~2 3e

21

,*

8 .~.g:

12 ~5

:. :.

.g:

m e 9 9 9 9d 9d9ddd 99 d dd d dd

a

a

aa

a .a .a-

.a .a a a a a a a a a a a a aaaaaa aasa3

? Y 2 * Yw

Y

E

.2.2.2-2 +; ;. 0 0 0 0 AAAA + A

.p 5.; E

+E

.2.2 .$k 3.8 Jg!

-2 ;. ~

+

A

A

.g

a.; A

a g . 0

.z

+;

AA A A ~A A ~P

j j j $3

MARINE ENGINEERING

ELECTRIC PLANTS

$f$

444

33s

d m -

0 0 0

aaa

A48

.333 zzz

e

J 3 b,

a

3

4 a p

UUU p"

PPP

a aaaa a a aa a az,o,o _o _o 23 3

23.

8~

ccc

-g

2s2

8z

e

Y Q

aaa 332

8

g E CI

%

PlPlPlPlPlPl 333333 ddddl4l4

g h a m 3333 d dl4l4l4

fY%flf

I lllf "0 %%%% ea ea ea ea

CCCCCC

0 0 0 0 0 0

ea*aOa eaeaeae

a"k g . . . .

aaaaaa aaaaaa 'd '6'd 'd 'd'd

-[

UUUUUU

da ssssss 0

N

8 C

30g

0

0

0

0

0

$C$C

Ndndnd

'C 'd '6.6 'C 'C 222232 N N N N N N

u du uu 0

0

0

0

0

%

ss ss

a

2s

.z .;.a

-0

2-

3.2

2 22 22

u uuuu 0

In

2

0

0

0

0

Lciuamo

t."4N

.z .g.g.g.g N N N N

a aaza

u u uuu u

U U . U U U U uuu uuu u

s s

zs

0

q d

0

0

0

0

$$@

0

$i

0

0

0

0

0

0

%Zg$

0

0

0

ZZs 9% 5

0

$ ?'?$ $ d ddn 4

'? t.8

1'?q

-odd

n d m

e$$ A d d

00

.a

.>2

e.z

2 2 2g.8

% c

.a

. -m nn

.a.z.a.a .a.a.; %.az g

2 $ 2222 2s.2 $ 2 $

a

washdown or possible transient submergence may be encountered. 5 Explosion-proof, fan-cooled-used in an atmosphere containing an explosive mixture. This type of enclosure is required to be capable of withstanding an explosion of a specified gas or vapor which may occur within it while preventing the ignition of the gas or vapor surrounding the enclosure. Explosion-proof equipment is not watertight and must be enclosed in a watertight housing if it is subject to weather conditions. The cooling fan should be constructed of nonsparking material and be protected by a guard. 6 Submersible, self-ventilated-used for those limited and special applications that may require normal operation in air and emergency operation when submerged. A positive means of providing the required capability is the use of a bell type of enclosure. The motor and pump combination is vertically mounted and covered with a close-fitting bell that is open a t the bottom. The bell must be of sufficient depth so that the required submergence will not force water onto the motor windings. Under special and restricted conditions established by the U.S. Coast Guard, submersible motor-driven pump assemblies may be used for pumping out cargo tanks of liquefied methane, propane, ammonia, and other unusual cargos, the primary restriction being that air must be excluded from the cargo tanks at all times so as to prevent the possibility of an explosive mixture existing within the tanks. 7 Dripproof protected encapsulated, self-ventilatedused for those applications that may be subject to temporary submergence, splashing, spraying, or hosedown. Witb encapsulated motors the windings around the end coils and in the slots are completely encased in a protective insulating coating that permits exposure to specified liquids, and the bearing cavities are made watertight. This type of enclosure is suitable for many applications that heretofore have required total enclosure. b. Terminal Boxes. All motors are normally furnished with terminal boxes having threaded pipe taps for ship's cable entrance terminal tubes. Motor terminal leads and ship's cable are mated by means of cable connectors within the terminal box. The degree of enclosure required for terminal boxes is usually the same as that provided for the motor. The desired location of terminal boxes and the number and size of tapped holes for cable entrance are normally specified by the shipbuilder. c. Insulation. Insulating materials for use in motors are divided into categories according to their ability to withstand high temperatures for long periods of time. These categories are Class A, B, F, or H. Class A insulation has the capability of operating a t a maximum temperature of 105 C; Class B is rated at 130 C; Class F a t 155 C; and Class H a t 180 C. These temperatures in each case represent the insulation syqtem material capability and are the summation of the ambient temperature, motor-winding temperature rise above

625

ambient, plus an estimated temperature gradient referred to as the hot-spot allowance. The regulatory bodies have specific requirements regarding temperature rise limits for the various classes of insulation for different applications. Bearings and bearing lubricants should be selected based on operating temperatures encountered with each class of insulation. d. Special Treatment. All motors for use in marine applications should be given a special impregnation to make the winqiigs resistant to salt water, salt air, oil fumes, and fungus. Metal parts are made of corrosionresisting materials or are treated to render them corrosionresistant. For example, small hardware is usually zinc-plated, and the shaft inside the frame, exposed laminations, and brackets are usually treated to prevent corrosion. Small hardware for motors exposed to the weather is normally stainless steel. e. Ambient Temperature. Motors for machinery spaces are designed for an ambient temperature of 50 C, normally using Class B insulation. Motors for use in areas other than machinery spaces are designed for an ambient temperature of 40 C, normally using Class A or B insulation. Exceptions involve unusually hot areas, such as a t the tops of boilers or adjacent to smoke uptakes, and certain axial-flow exhaust fan applications for which the motors are rated for 65 C using Class F or H insulation. The trend is to use Class F insulation for both 40 C and 50 C spaces, to reduce the overall size of the motors. f. Space Heaters. Motors.subject to wide variations of temperature or excessive moisture conditions are often provided with space heaters to prevent moisture condensation in the motor when idle. The heaters may be resistance units bolted to the inside of the lower frame or a phase winding energized through a low-voltage transformer. I n either case, the heating circuit is electrically interlocked so as to remove heater power whenever the motor is energized. g. Shafts. For flexible and rigidly coupled drives, National Electric Manufacturers Association (NEMA) standard short-shaft extensions are used. However, in some instances shafts are provided with the end tapered, threaded, and equipped with a nut and washer for ease in disassembly and reassembly. Brake mqtors or motors using shoe brakes are provided with front-end shaft extensions as required by the brake application. NEMA long-shaft extensions are provided for pulley-driven auxiliaries. Special long shafts are provided for impellers of close-coupled pumps and axialflow fans. Carbon steel sha@ are normally provided for coupled drives and for freshwater closecoupled pumps. Shafts made of corrosion-resistant materials (e.g., stainless steel or monel) and fitted with sle$*s are generally required for pumps handling corrosive liquids. h. Bearings. With few exceptions a-c and d-c motors are equipped with greasable ball bearings; axialflow ventilation fan motors are usually equipped with prelubricated sealed ball bearings, since they are located in ductwork and not easily accessible. Sleeve bearings,

i II I

I

!

I

1

1

626

-

TORQUE (PER CENT OF FULL L O A ~ T O R Q U E ~ -

TORQUE (PER CENT OF F U L L - L O A ~TORQUE)

I

I

Fig.

ELECTRIC PLANTS

MARINE ENGINEERING

7 Torquespeed curves for induction motors rated 30 to 50 hp

designed for flood lubrication or forced lubrication, are used only in special applications. i. Speed Reduction Drives. Separately mounted coupled reduction gears or integrally mounted reduction gears are used to decrease or increase the speed obtained from the drive motor. Motors and gears must be carefully selected to suit the speed, torque, and duty of the driven auxiliary. A V-belt drive is often used for air and refrigerating compressors and slow-speed fans. The use of a V-belt drive requires the motor to be mounted on a base that affords adjustment of the belt tension. j. Mounting. Motors are designed for mounting in any required position, i.e., horizontal, inverted horizontal, vertical with the drive shaft up or down, and in some instances inclined. Most shipboard auxiliaries are driven by horieontal motors mounted on a common bedplate with the driven machine. However, the use of vertically mounted motor-driven centrifugal pumps provides a saving in deck area and a preferred piping arrangement. In addition to general-purpose mountings, NEMA has standardized two end mounts, types C and D, and two flange mounts, types P and PH. Types C and D are used either horizontally or vertically, with the relative location of the face and .feet fixed by the standards. Types P and PH flanges are used for vertical pump applications. Each type of mounting should be coordinated with the driven auxiliary to insure a satisfactory fit and performance. For heavy assemblies (such as an overhung motor with a disk brake attached by a flange), it is advisable to provide a foot-mounted motor to afford rigidity. 4.3 iElechical Characteristics of Alternating-Current Motors. The speed, horsepower, and duty rating of a

motor is fixed by the required input to, and the operating cycle of, the driven machine. Duty ratings. (operating cycle) for shipboard applications are classified as continuous duty or intermittent duty. Continuous duty is a requirement of service that demands operation a t a

Fig. 6

---

Design D Induction motor torque-speed curve for various slip values

substantially constant load for an indefinitely long time. Intermittent duty is a requirement of service that d e mands operation for alternate periods of (a) load and no load; (b) load and rest; or (c) load, no load, and rest, with such alternate intervals being definitely specified. The applicable IEEE Standard No. 45 contains a description of duty ratings for specific applications. The majority of the motors used aboard ship are squirrel cage induction motors. The design designations, Design A, B, C, and Dl for three-phase squirrel cage induction motors, are based on torque, current, and speed requirements. Each design offersdifferent torque, speed, and current characteristics to meet various operating requirements as may be seen from Figs. 7 and 8. Design A motors have a normal starting torque, high starting current, and low slip. This motor is not used for the usual shipboard applications because of its high starting current characteristic. Design B motors have a normal starting torque, low starting current, and low slip;. This is the motor most commonly used on shipboard; it is generally used for centrifugal pumps, fans, blowers, motor-generator sets, and compressors that are not loaded when started. Design C motors have a high starting torque, low starting current, and low slip. This motor is normally used for applications such as steering gear, anchor windlass, plunger-type pumps, and compressors that are not unloaded when started. Design D motors have a high starting torque, moderate starting current, and high slip; this motor is normally used for capstans, winches, valve operators, conveyors, elevators, and hoists. The squirrel cage motor may be designed for one, two, three, or four speeds. In addition, this motor is suitable for adjustablespeed operation when used with a combined frequency and voltage control; adjustable speed by voltage control only requires a specially designed highslip motor. Wound-rotor motors may be used for adjustablespeed

or constant-speed applications. For adjustablespeed applications, up to 50-percent speed reduction can be obtained by inserting different values of resistance in the rotor circuit by means of multistep controllers. Each set of resistance values inserted in the rotor circuit results in different torque and speed characteristics as shown in Fig. 9. The use of wound-rotor motors for constant-speed applications is limited on shipboard; however, its use should be considered when it is necessary to start a large motor from a relatively low-capacity generator. The advantage in this case would be the low starting current and normal starting torque that are available with proper selection of secondary resistances in the rotor circuit. Typical applications requiring adjustable-speed selection on shipboard are forced-draft blowers, fuel-oil service pumps, and main circulating pumps. Although wound-rotor motors may be used, multispeed squirrel cage motors are more commonly used for these applications. , Synchronous motors are used in shore practice for improving the system power factor and are usually applied to drive large continuous loads such as motorgenerator sets, compressors, pumps, and fans. They have practically no application on shipboard but may be used for such applications as large-capacity circulating pumps and motor-generator sets to improve the system power factor.

2EROEXTERNAL RESISTANCE 2-4 CENT SLIP / PER AT FULL LOAD

Fig. 9

Typical torque-speed curves for a wound-rotor motor with various external resistances

loads with lower armature current and better commutation than with a shunt winding only. The compound winding is desirable for loads of high inertia such as certain direct-connected centrifugal fans, as well as for such shipboard applications as the propeller shaft 4.4 Electrical Characteristics of Direct-Current Motors. turning gear, valve operators, compressors without The major distinctive characteristic of d-c motors is unloaders, and positive-displacement pumps. The crane or winch type of winding is a compound their type of field windings. Direct-current motors may winding consisting of a light shunt field and a heavy have a shunt winding, stabilized-shunt winding, series series field. These motors are deigned for a specific winding, cornPofind winding, or a crane or winch type of load and afford many desirable characteristics such EM winding. Shunt or stabilized-shunt windings are used for high torque for heavy loads and high speed for light applications requiring constant speed rkgardless of load loads with the light shunt field providing constant-speed variation. Typical shipboard applications are fans, characteristics. Typical shipboard applications of this type of winding are cargo winches and anchor windlasses. blowers, centrifugal pumps, and elevators. Motors with stabilized-shunt windings are provided 4.5 Mechanical Characteristics of Motor Control with a light series field in addition to the shunt field to Equipment prevent a rise in speed as the load increases; hence the a. Starter and Controller Panels. A grouping of term "stabilized". This type of winding is desirable for loads of high inertia such as direct-connected centrifugal several motor starters housed in a free-standing deckfans or pumps, since acceleration is accomplished with a mounted structure is known as a group control or motor lower armature current and better commutation than control center. Each motor starter within a group control is energized from a common power supply feeder with a common shunt winding. Series-winding motors are used for loads requiring a through individual circuif, breakers mounted in each very high starting torque. They are also used in starter. Group controls &-metal-enclosed unjts having applications requiring operation over a wide speed range metal barriers between starters. In general, individual starter enclosures are either such that the motor develops a high-speed operation at light loads, and at low speeds a comparatively light dripproof, watertight, submersible, or explosion-proof as current- and a high torque. Series motors are partic- required by their location. Starters mounted in group ularly suitable for the operation of warping capstans control switchboards are generally of the "open" type which require a high torque a t low speeds for handling mounted in a dripproof enclosure. Cabinet enclosures are usually designed for bulkhead loads and a high speed at light load for retrieving lines. The friction loss of the machinery is generally sufficient mounting in smaller sizes and for deck mounting in weights over 150 lb. Consideration should be givenib , to limit the no-load speed to a safe value. Motors with compound windings develop a high the need for rear access on large starter panels for starting torque and have the ability to handle peak connections and inspection.

1 628

MARINE ENGINEERING

ELECTRIC PLANTS

i

Suitable provisions for cable entrance are made in all cabinets, usually a t the top for the feeder and at the bottom for the motor and control leads. Watertight terminal tubes or other adequate seals to exclude moisture should be provided where cables enter at the top of dripproof cabinets. Cable clamps or other adequate means for holding cables in place to prevent chaffing from vibration should be provided where cables enter the bottom of dripproof cabinets. All cable entrances in watertight starters are made through terminal tubes. In general, practically all motor starter enclosures, excepting explosion-proof starters, are provided with removable drilling plates so as to avoid the danger of metal chips dropping in the working parts as could occur if drilled after delivery. Explosion-proof enclosures are often purchased with the number and size of tapped holes for cable entrance terminal tubes specified. Separately mounted resistors, if required, should be installed in well-ventilated spaces and provided with protection from dripping liquids. The ambient temperature and class of insulation for motor control equipment are determined in the same manner as previously described for motors. b. Master Switches. The broad category of master switches includes pushbutton, drum switches, selector switches, pressure switches, temperature switches, interloclr sw~itches,float switches, and any other type of remote pilot device required for proper operation of the controlled motor through starter equipment. Local switches and indicating lights which form a part of controllers should have the same degree of enclosure as the controller. For remote locations, the enclosure of switches should be dripproof, watertight, submersible, or explosion-proof to suit the requirements of the location. In general, for weather or corrosive areas, watertight enclosures of cast bronze or brass for small switches and bronze, nodular iron, or stainless steel for winch control switches are provided. The cable entrance into master switches should be through a bushing in the bottom of dripproof switches and through watertight terminal tubes in all others. The regulatory body rules generally require that controllers be mounted adjacent to the driven auxiliary. With group control and in those situations where the controller cannot be mounted in sight (as for certain fans and forced-draft blowers), remote pushbuttons, designed so that the stop button $an be maintained open, are located a t the driven auxiliary. c. Speed-Regulating Rheostats. For shipboard use, speed-regulating rheostats for shunt field control of direct-current motors are provided with dripproof covers and enclosed terminals and are arranged for bulkhead mounting. This construction permits a convenient location of the rheostats with a minimum probability of damage to the rheostat or injuries to personnel. 4.6 Electrical Characteristics of Alternating-Current Motor Controllers., Motor controllers are designed to

perform definite electrical functions regarding the control and protection of motors. The characteristics and

applications of the various types of controllers that provide these functions are as follows: a. Controller Operation. Controllers are designated as being either manual or magnetic in operation: Manual controllers are normally used for applications of less than 2 hp that require only "on-off" operation. Magnetic controllers are used for all other applications and may be classified as automatic or nonautomatic. Automatic controllers start and stop the motor in response to some controlled factor with no attention from an operator. Nonautomatic controllers require manual operation of a pushbutton or switch to initiate a start or stop. After the initial manual operation, the controller completes the starting or stopping of the motor. Nonautomatic controllers are used for practically all applications other than those auxiliaries that require automatic cycling such as air compressors and refrigeration compressors. Typical control-circuits for a-c magnetic controllers are shown in Figs. 10 and 11. b. Types of Controllers. Controllers of the acrossline type are used for practically all shipboard auxiliaries since the ship's generating plants are usually of adequate capacity to handle the starting currents of all motors installed. However, controllers of the autotransformer type are used when it is necessary to limit the starting current of a motor so as to avoid imposing an excessive load on the generating plant. This type of controller should be designed for closed circuit transition so as to avoid high transition currents. Standard starting transformers for motors above 50 hp have taps of 50, 65, and 80 percent of full voltage; only 65- and 80-percent taps are provided in sizes below 50 hp. The starting current drawn from the line is proportional to the square of the percent voltage tap; i.e., 80 percent tap equals 64 percent of the across-line starting current. A typical application is for fire pumps that must be started from relatively small-capacity emergency generators. A primary resistor type of controller could be used to limit the starting currents of large motors and also for speed control of small motors; however, their application is very limited because the motor starting current is not substantially reduced and therefore they provide little or no advantage for shipboard use. Controllers of the secondary resistor type are used to limit the starting currents and provide speed control for wound-rotor induction motors; typical applications are forced-draft blowers and main circulating pumps. A wye-delta type of controller could be used to limit the starting currents of large motors. With this arrangement the motor is started in the wye connection and then reconnected, with closed transition, to the delta running connection. This arrangement requires a six-lead motor and the starting current would be approximately 33 percent of the across-line starting current. A typical application is a bow thruster. See Fig. 12 for typical methods of starting a-c induction motors. c. Controller Protective Features. Low Voltage

4

I

4

1 t

f I

1

-

,I

Protection (LVP) is the feature that is provided to cause the controller to disconnect the motor from the power supply upon reduction or loss of voltage; the motor remains disconnected until the voltage is restored and the motor is restarted by manual operation of its starting pushbutton, as shown in Fig. 13. This feature is provided as a means of preventing the simultaneous restarting (after an interruption of the supply voltage) of a large number of motors such that their large starting currents are additive. Low' Voltage Release (LVR) is the feature that is provided to cause the controller to disconnect the motor from the power supply upon a reduction or loss of voltage; the motor remains disconnected until the voltage returns, .and then automatically reconnects the motor to the power supply to restart it, as shown in Fig. 14. This feature is usually applied to only those vital auxiliaries that must be automatically restarted immediately upon restoration of power. Typical examples are lube-oil service pumps, main and auxiliary condensate pumps, main circulating pumps, and control air compressors. It is usually desirable to use timedelay relays with LVR controllers to obtain staggered starting and prevent simultaneous restarting of all LVR auxiliaries. Motors that are automatically controlled by pressure-switches, and similar devices, have an inherent LVR feature unless a low-voltage relay, which opens upon failure of line voltage, is provided in the control circuit. All manual-type controllers provide LVR characteristics. Overload protection is the feature that results in the controller operating to disconnect the motor from the power source when excessive currents (not short circuits) occur that could cause overheating of the motor. This feature is provided by overload relays; separate relays are required for each winding of multispeed motors. Overload relays may be either thermal or magnetic. Thermal overload relays generally consist of a heatsensitive element and a heat-generating 'element. The heat-generating element may be a heater or coil in series with the motor load circuit. An excessive motor current passing through the heat-generating element causes the heat-sensitive element to react to open the overload relay contacts, thus breaking the circuit to the operating coil of the main line contactors; this in turn causes the contactor to open the motor circuit. Since the tripping characteristics of the thermal overload relay depend on both the length of time of application and the amount of overload current, the relay can be, and normally is, designed to follow approximately the timecurrent heating curve of the motor. This curve represents the values of the current that a motor can carry for different lengths of time without damaging the motor insulation. Thermal overload relays should be compensated against possible ambient temperature changes. Thermal overload relays are generally of the solder-pot, bimetal, single metal, or induction type. Practically all applications of overlaad relays on shipboard are of the thermal type.

Magnetic overload relays generally consist of a coil in series with the motor load circuit and a tripping armature or plunger. When the amount of overload current for which the relay is set passes through the series coil, the tripping armature is actuated to open the overload relay contacts, thus breaking the circuit to the operating coil of the main line contactors; Bhis in turn causes the contactor to open the motor circuit. Magnetic overload relays are not affected by variations in the ambient temperature and &quire no temperature compensation. Magnetic overlo*ad relays are of the instantaneous or time-delay type and have limited application on shipboard because they do not use heat in their operation and consequently do not follow the heating curve of motors. Overload relays are provided with a means of resetting so that the motor controlled can be restarted with overload protection. Tripped thermal overload relays must be allowed to cool before the tripping mechanism can be reset. Magnetic overload relays can be reset immediately after tripping. The three forms of overload relay resets are manual (hand), automatic, and electric. The manual form is the most common for shipboard use and consists of a rod or lever which, when operated, causes the tripping mechanism to be returned to its original position. The automatic reset has no practical application on ships. The electrical reset is limited to those applications where it is desirable to reset a relay from a remote operating position. Emergency run features should be provided on controllers for certain auxiliaries, such as elevators, in which case stopping in the middle of an operating cycle could be highly undesirable. This feature, which is initiated by operating a pushbutton or lever, renders the overload relay tripping mechanism inoperative so that the auxiliary can be operated with the motor running in an overload condition until the operating cycle is completed. In general, each control wire that leaves a controller should be provided with short-circuit protection. Such protection may be provided by a fuse, located in the controller, if the lead is not already protected by a current-limiting device (coil or resistor) located i11 the enclosure. When a secondary source of power greater than 24 volts is brought into a motor controller for alarm, indicating light, or other circuits, a suitable interlock is usually provided to disconnect the secondary source upon opening of the controller door. In lieu of disconnecting the secondary source as noted in the foregoing, an independent disconnect deyice may be used for this purpose. This independent CP-iconnectihould be-located adjacent to the motor and controller disconnect, and a sign should be provided on the main disconnect to warn that both devices should be operated to disconnect completely the motor and controller. 4.7 Electrical Characteristics of Direct-Current Motor StaHers. Direct-current motor controllers may be of

either the across-line or resistor type. Full-voltage across-line controllers are usually manually operated and

MARINE ENGINEERING

630 LINES L2

LI

ELECTRIC PLANTS LlNES L t L21 L31 TDS TDS TDS

L3 &

1

& A - - 4--ADS

I

H, :I: a

ELEMENTARY DIAGRAM II M OL II

, , OL N.

TI

N

T2

MOTOR

I

H2

CR 26

L(LA 80

*

PS M

(

UNLOADER SOLENOID

(L) LOCAL (R) REMOTE

DESCRIPTION OF OPERATION

ELECTOR IN "HA ": TO START. PRESS A START BUTTON ENERGIZING SELF MAINTAINING RELAY "cR': THE "CR" E o N T A c T s CLosEYNEasIzING MAIN CONTACTOR "M? THE * ~ ~ ~ c o r y CONNECT A c n THE MOTOR ACROSS THE LINE AND ENERGIZE.7HE UNLOADER SOLENOID. TO STOP PRESS A STOP BUTTON REMOVING OCR" "M"AND THE MOTOR FROM THE LINE. SELECTOR AUTO^: PLACING THE SELECTOR s w ~ i c n IN AUTO^ INSERTS A N.B. PRESSURE SWITCH, *PS*: IN

~k

Fig. 11 Typical elementary and wiring d l a g r m for a potable water pump application

...- --.. ...----,..

THC r n N T D n 1 I CD

TO START PRESS A START BUTTON ENERGIZING RELAY "CR". WHEN THE "PS~CONTACTSCLOSE, MAIN CONTACTOR "M" IS ENERGIZED CONNECTING THE MOTOR ACROSS THE LINE. WHEN THE n ~ "Ma'IS DROPPED REMOVING THE MOTOR FROM TUC I INC TO STOP. PRESS A STOP BUTTON R E I I N AVOLTAGE FAILURE CONDITION, A L L vv~la PRESS A START BUTTON AFTER RESTORATwn vr r v ~ ~ r r utc~. v n IN AN OVERLOAD CONDITION, THE 0.L.CONTACTS OPEN REMOVING THE RESTART, PRESS A RESET AND THEN A START BUTTON. (OVERLOAD PROTECTIOIY

~

w

~

~

~

Ra. 10 Tvdcal e l e m ~ t aand ~

I.

~

~

~

~

~

~

~

~

TO START THE MOTOR MANUALLY, TURN THE SELECTOR SWITCH TO THE MAN POSITION. THIS ESTABLISHES A CIRCUIT TO THE MAIN CONTACTOR M. THE MOTOR WlLL START AND CONTINUE TO RUN AS LONG AS THE SELECTOR SWITCH IS IN THE MAN POSITION. TOOPERATE THE MOTOR AUTOMATICALLY UNDER THE CONTROL OF THE PRESSURE SWITCH PS. TURN E SWITCH W ~ L LKENE ESTABLISH A-~IREUIT THE SELECTOR-SWI~CHTO THE AUTO POSITYON.Y ~ PR&RE TO THE MAIN CONTACTOR M. THE MOTOR WlLL START AND CONTINUE TO RUN UNTIL THE PRESSURE SWITCH CONTACTS OPEN. THE MOTOR WlLL THEN CYCLE OFF AND ON AS THE PRESSURE SWITCH CONTACTS OPEN AND CLOSE. AVOLTAGE FAILURE WlLL CAUSE THE MAIN CONTACTOR M TO OPEN AND DISCONNECT THE MOTOR FROM THE LINE. WHEN VOLTAGE IS RESTORED, M WILL RECLOSE AND START THE MOTOR (LOW VOLTAGE RELEASE). AN OVERLOAD WILL CAUSE THE OL CONTACTS TO OPEN, DEENERGIZING M AND STOPPING THE MOTOR. TO RESTART. PRESS THE RESET BUTTON. THE MOTOR MAY B E STOPPED AT ANY TIME BY TU8NING THE SELECTOR SWITCH TO THE STOP POSITION. ~

LINES

,

REAR VIEW OF DOOR

FRONT VIEW OF PANEL

ITO

LINES REAR VlEW OF DOOR

, 1-

FRONT VlEW OF PANEL

limited to those applications involving a maximum of 2 hp subject to the driven machinery withstanding the resulting high starting torque and the motor handling the high inrush current. Good practice restricts this type of starting to small galley appliances, small pumps, and ventilation fans. Controllers of the resistor type are used for most d-c motor control apylicatibns. They provide mean8 of

reducing the motor starting current to prevent motor damage and prevent overloading the generating plant. The resistor type of controllers may be either manual or magnetic. Most controllers of the manual d-c resistor type are similar to acros9-line controllers except for the addition of a starting resistor and an accelerating contactor. The control circuitry is designed such that as the motor comes up to speed the accelerating con-

tactor is actuated to by-pass the starting resistance and connect the motor to full voltage. Manually operated "dial" or "face plate" starting rheostats may be used for starting and speed regulation of nonreversing motors. They are used for light-starting auxiliaries such as ventilation fans but should be limited to ratings below 2 hp. Practically all d-c controllers are of the magnetic, reflistor type. Magnetic controllers may be of either the nonautomatic or automatic type and consist of a main contactor, control relays as required, one or more accelerating contactors, an overload relay, emergency run feature, and starting resistance. The main contactar is controlled by a manually operated master

switch (nonautomatic control) or by an automatically operated device such as a pressure-regulating switch (automatic control). When the main contactor closes, the motor is connected to the power supply through the starting resistances. The p n t r o l ciquitry contains current-sensitive series relays and accelerating contactors so that as the motor comes up to speed, individual steps of the starting resistance are by-passed until all steps are out of the motor circuit and the motor is connected across the line. Another method of control starting uses definite time accelerating contactors to close automatically and shunt out steps of the starting resistance. Speed adjustment for d-c motors may be obtained by

MARINE ENGINEERING

ELECTRIC PLANTS

REDUCED-VOLTAGE AUTO TRANSFORMER

LINES

I

633

ELEMENTARY DIAGRAM

START

DESCRIPTION OF OPERATION

Fig. 13 Typical motm canhol circuit with low-voltage prohctim feature

(01 ACROSS LINES

TO S V R T THE MOTOR PRESS TYE START BUTTON. THIS ENERGIZES CONTACTOR M. CONNECTING THE MOTOR ACROSS THE LINE. M MAINTAINS ITSELF THRU ITS OWN AUXILIARY CONTACTS. TO STOP THE MOTOR, PRESS THE STOP BUTTON. A VOLTAGE FAILURE WlLL CAUSE M TO OPEN STOPPING THE MOTOR. TO RESTAkT, WHEN VOLTAGE IS RESTORED, PRESS THE START BUTTON. (LOW VOLTAGE PROTECTION). AN OVERLOAD WlLL CAUSE THE OL CONTACTS TO OPEN CAUSING M TO OPEN STOPPING THEMOTOR. TORESTART.PRESSTHERESETBUTTONANDTHENTHESTARTBUTT0N.

(b) CLOSED TRANSITION LINES

SERIES

CLOSED- CIRCUIT TRANSITION

Fig. 12 Typical mehds of starting a-c induction moton

IM WOUND-ROTOR INDUCTION MOTOR

IA a Z A ARE DELAYED IN CLOSING BY A PRESET TIME

R I CONTACTOR SEO.

(C) REDUCED-VOLTAGE

PRIMARY RESISTOR

MOTOR

(dl FULL VOLTAGE SECONDARY RESISTOR

using a rheostat in the motor shunt field circuit or by varying the amount of resistance of continuous-duty resistors in the motor armature circuit. 4.8 Brakes. Electric brakes for marine service are of either the disk or shoe type, each type being spring set and magnetically (solenoid) released. Solenoids or magnets are usually energized (brakes released) through contactors whenever the associated motor is in opera-

tion. Deenergizing the motor also deenergizes and engages the brake. Disk brakes may be either a-c or d-c operated; shoe brakes are usually d-c operated. For motors that are 50 hp and larger, d-c operated brakes are usually provided. I n general, brakes for suspended loads are rated a t 200 percent of the motor torque. For other types of loads,

the brakes are rated to stop the load under any operating condition involved; in these cases, the brakes are usually rated a t 100 percent of the motor torque. Brakes located in weather locations should be of a watertight construction and have electric heaters to prevent the accumulation of condensation during nonenergized periods. Brakes should be provided with a means of being mechanically released for emergency operation in the event of a power failure. Disk brakes are attached directly to the motor fronb end bracket and require no special foundation; from a space point of view, disk brakes lend themselves to an

economical installation. Most shoe brakes are foob mounted and require a special foundation for proper alignment with their motor. 4.9 Transformers. Transformers _ are utilied to supply alternating-currefif* loads which cannot be operated on the ship's primary voltage. Transfortners are normally of a single-phase, 60-cycle, air-cooled, dry type that is designed for continuous duty with dripproof enclosures and suitable for connecting in a threephase bank. Each transformer in a bank should have an identical rating with the primaries connected in delta and the secondaries connected in delta or wye as re. quired for the intended service.

634

MARINE ENGINEERING LINES LI

L2

635

ELECTRIC PLANTS

ELEMENTARY DIAGRAM

lead-acid or alkaline type. It is advisable that all storage batteries provided on a given vessel be of the same type, to prevent the possibility of contaminating any battery electrolyte through the inadvertent use of a common hydrometer. Principal applications for storage batteries are: (a) emergency power for radios; (b) no-break power supply, (c) diesel-generator cranking; (d) lifeboat engine starting;

L3

IDS IDS IDS MOTOR OL

(e) emergency lighting when an emergency diesel generator is not installed; (f) fire-screen doors and watertight doors; (g) forklift trucks; (h) general alarm system; and (i) certain interior communication loads. The regulatory bodies h a w published specific detailed requirements regarding the construction, rating (capacity), ventilation, installation, and arrangement of i storage batteries.

OFF

-RUN

I

Section 5 Lighting Fixtuns and Equipment

2

DESCRIPTION OF OPERATION TO START THE MOTOR POSITION THE SELECTOR SWITCH &T"RUN". THIS ENERGIZES CONTACTOR M CONNECTING THE MOTOR ACROSS THE LINE. TO STOP THE MOTOR POSITION THE SELECTOR SWITCH AT"OFF~. A VOLTAGE FAILURE WILLFAUSE M TO OPEN STOPPING THE MOTOR. WHEN VOLTAGE IS RESTORED, THE MOTOR WILL IMMEDIATELY RESTART (LOW VOLTAGE RELEASE). AN OVERLOAD WILL CAUSE THE OL CbNTACTS TO OPEN CAUSING M TO OPEN STOPPING THE MOTOR. TO RESTART, PRESS THE RESET BUTTON.

LINES REAR VIEW OF DOOR

Li L21 T" T T

VIEW OF PAWL

L3f

I

.

,

Fig. 14 Typical motor m t r d cirwl with law-valtage please feature

4

I t

I I

RESET BUTTON PUSH TO RESET

u

MOTOR

06

'

The kva ratings of transformers should be based the connected load plus a reasonable allowance for installed spares and future development. Every effort should be made to balance the load between phases of transformer banks. 4.10 Motor Generators. Motor generators are provided to supply power to loads requiring special voltages and/or frequencies that differ from the ship's primary power. The principal applications for motor-generator sets are: (a) Providing d-c power for cargo handling equipment; (b) Providing a-c power for the "temporary emergency" loads on passenger vessels; in this applic* tion, upon loss of normal power, the motor-generator set

is supplied with power from an emergency battery and in turn provides a-c power to selected vital emergency loads until the emergency diesel-generator can start and assume all emergency loads; and (c) Providing power for automatic elevators on passenger vessels. 4.1 1 Reetiflers. Rectifiers are provided to supply loads requiring d-c power that is not available from the ship's primary power. Principal applications for rectifiers are: (a) Electronic equipment requiring d-c power; (b) Battery charging; and (c) Fire-screen doors. Specific details regarding requirements for rectifiers are contained in IEEE Standard No. 45. 4.12 Batteries. Storage batteries are usually of a

i

5.1 General. Marine fixtures and appliances must be of a special design and construction to suit the various requirements incidental to shipboard installations. They must be rugged to withstand normal shipboard vibration. Furthermore, enclosures must be compatible with the particular environment in which they are located; also, they must be of a corrosionresisting material or have an effective corrosion-resisting finish. To prevent rapid deterioration of the finish, decorative fixtures with polished or plated surfaces must be protected from the salt atmosphere by some form of protective coating, such as lacquer. The possibility of radio interference from fluorescent or high-intensity electric-discharge lamps also requires consideration in the selection of light sources. All lighting fixtures, with few exceptions, are required to meet the standards of the Underwriters Laboratories, Inc., publication UL595, Marine Type Electric Lighting Fixtures. Fixtures not covered by this standard must have U.S. Coast Guard approval for each specific application and vessel. In developing the design of the lighting system, the "Recommended Practice for Marine Lighting, " published by the Illuminating Engineering Society, should be used for determining the minimum footcandle requirements. 5.2 Types of Fixtures. Fixtures may be classified by the types of lamps used. Fluorescent, incandescent, and high-intensity electric-discharge lamps are the main types of lamps used. Of the three, the fluorescent offers the most advantages and is replacing the incandescent for practically all general illumination requirements. The color rendition of fluorescent lamps has been steadily improved in recent years and is now almost the equal of the incandescent lamp in this respect. Their high luminous &cacy makes them much better suited for use in air-conditioned spaces due to lower heat generation; and their longer life considerably reduces maintenance. The incandescent lamp provides a compact, highbrightness source of light which can easily be directed by a small, simple luminaire. This feature makes the incandeacent lamp superior for spot and detail illurnin* tion. It operates readily on either &C or d-c supplies. It is less sensitive to ambient temperatures than other light sources and, consequently, is preferred where

illumination is required in severe hot or cold temperatures, i.e., refrigerated spaces, weather locations, detail lighting for boiler gages, etc. Incandescent lamps are the most practical for minimal illumination requirements, such as small locker rooms and indicator lights. The high-intensity electric-discharge lamps, in general, are compact, with high brightness sources and have an even higher luminous efficacy and a longer l i e than the fluorescent. %lost of these lamps have an acceptable color of light-----.__ W w o r k areas, such-aScZF6 KoIds,"aeEk -' fEGIh$tTng, and engine~ooni~;but-arenot consiaered-~ ~ e ~ T o ~ 5 E 5 a T l J m m m - ~ ~ M color -rendition. .Their opertttion 'and maintenance ecGmYS-eGen more pronounced than for the fluore* . cent. Their long lamp -life, ranmgto 24;MlU-%r and 45i%t,er1 makes them an ideal lamp for fixtures mounted high on masts and kingposts and other locations not easily accessible. The problem of relamping is reduced such that the maintenance for this type of fixture is practically negligible. Of the high-intensity electricdischarge lamps, the mercury-vapor is most acceptable for shipboard use. For certain applications, supplementary instantaneous lighting may be necessary since, in the event of power failure, instantaneous relighting is not accomplished by electric-discharge lamps. The use of some of these lamps is limited due to certain hazardous characteristics (lamps containing sodium may cause a fire when broken on a wet surface). Serious corrosion condition8 may also result if the mercury from a ruptured mercury-vapor lamp comes into contact with aluminum. Even with these limitations, however, the mercury-vapor lamp is finding its rightful place on shipboard. Lighting fixtures may glso be identified according to their application; i.e., cedi% lights (located overhead in ceiled spaces), deck fixtures (located overhead in unceiled spaces), bulkhead Mures, detail lights (desk, berth, mirror, etc.), floodlights, navigational lights, and I miscellaneous lighting fixture4 --'%Sa*-faclr baiic types of fixture enclosures for shipboard use: watertight, dripproof, explosion-proof and nonwatertighi--Waterti@itTiiiiipment is installd b r d - ~ p t t . ~ e ~ Wwould i ~ i tbe exposed to seas, splashing, or severe moisture conditions, and for all installations in the weather. Dripproof equipment is installed in the % _

MARINE ENGINEERING

overhead in other wet or damp locations. Explosionproof equipment is installed in all areas that are subject to volatile gas accumulations; this equipment is so constructed that an explosion of a specified vapor within the equipment will not cause the ignition of the volatile atmosphere surrounding it. Nonwatertight equipment is used where moisture and volatile gases are not problems. 5.3 Passenger Stateroom Lighting. General illumination is provided for passenger staterooms by direct or indirect lighting supplemented by b m l i g h t s , mirror lights, table lamps, desk lights, bracket lamps, etc., to s G t 3 h e =gement %fTGmiture and decorative scheme. The trend is to the use of fluorescent coves, which may be combined with curtain valance lighting and incandescent down lights. A berth light, either separate or incorporated in a table lamp, is provided a t the head of each berth or bed. It should provide adequate light for reading but avoid annoyance to occupants of other berths in the room. Illumination for grooming at the dresser may be provided by bracket lamps on each side of the mirror, table lamps on each end of dressers, or by a mirror light recessed in the ceiling. These lights should be located or arranged for best illumination of the person seated at the dresser. I n bathrooms, general illumination is provided by one or more ceiling fixtures, with one or more mirror lights provided for each mirror. In small baths, where the mirror light provides the required level of general illumination, the ceiling lights are omitted. A receptacle is provided at each dressing table and a t each lavatory for the convenience of the passengers. 5.4 Officers' and Crew Quarters. Ceiling fixtures in the messrooms, lounges, recreation rooms, offices, and staterooms of the officersand crew usually are fluorescent with diffusing lenses. Berth, lavatory, table lamps, and desk lights are provided to suit furniture arrangements. Receptacles are installed at the lavatories, desks, and also for bracket fans when required. The toilet and shower spaces are illuminated by a ceiling light and a mirror light at the toilet case. I n small spaces where the mirror lights will provide the required illumination, the ceiling lights are omitted. 5.5 Passage Lighting. I n ceiled passageways, fixtures are usually fluorescent of the cornice type with concealed wiring. In unceiled passages, a similar type of fixture located on the bulkhead below interferences and angled to illuminate the center of the passage should be used. Passage lights are spaced from 7 to 9 f t apart and a light iii-IoCded'at the-interie6tlrdn+ofathwar2ship and fore-and-aft passages. Receptacles spaced about 40 f t are provided in passages for the attachment of vacuum cleaners and floor polishers. 5.6 Public Space Lighting. Architects and interior decorators generally design the passenger spaces, including the selection of furniture and fittings. They also determine the general character of the lighting installation and the type of fixtures to be used. The electrical

designer is responsible for the detail $esign of fixtures, as desired by the decorator, in accordance with sound engineering principles and for their suitability for marine installation with respect to rigidity, accessibility for servicing, ample ventilation, and adequate intensity of illumination. For illumination of the lounges, smoking room, dining saloon, and similar spaces, including adjacent foyers and passages, established practice favors simplicity with dependence on murals for decoration and with totally indirect trough or cove lighting arranged to provide adequate general illumination and to accentuate the decorative features. If direct illumination is provided, the detail design is determined by the fixture manufacturer's specialists in interior decoration, but all fixtures must be checked carefully by the electrical designer for compliance with marine practice. Special fixtures& public spaces include desk lights, table lamps, and floor lamps which are usually provided with built-in switches. Illuminated signs are provided on each deck a t passenger stairways and elevators for identification of public spaces, and for guidance of passengers to the debarkation decks in event of an emergency. Illuminated signs also are provided for all public toilets. When a toilet entrance is located in an athwartship passage, an additional sign is installed at each end in the main fore and-aft passages. Receptacles are provided, when desired, in floors or baseboards for lamps and vacuum cleaners. Depending on conditions, floor receptacles may be of the shallow type with the box for the fitting and conduit for wiring imbedded in the plastic floor covering, or they may be of the deep type with the box extending through the deck for wiring below. 5.7 Commissary Space Lighting. Careful consideration should be given to the illumination required for the preparation and inspection of food. General illumination for galleys, pantries, and service areas is provided by overhead flush-mounted dripproof fluorescent fixtures. Incandescent lights are installed under the exhaust hoods for the illumination of the ranges, griddles, etc. These fixtures must be constructed to withstand the intense heat from the range tops. Refrigerated spaces are lighted by guarded watertight deck fixtures controlled by a switch located within the space near the door. A fixture with a red globe, wired in parallel with the refrigerated space lights, is installed outside the door to show when the lights inside the compartment are "on". 5.8 Hospital Lighting. General and detail illumination for the hospital space is installed similar to a crew stateroom. For vessels having an operating room, a special operating light (or lights) is provided over the operating table; and all electrical equipment installed less than five feet above the deck must be explosion-proof. The Coast Guard rules should be reviewed to determine additional requirements as may be applicable to operating room installatibns. 5.9 Workshop Lighting. Illumination should be

637

ELECTRIC PLANTS

concentrated over the workbenches, machines or other areas where close visual tasks are performed. The lighting fixtures should be so located that the work area is not shadowed by the workman. Incandescent machine tool lights are also provided on lathes, drills, etc. Where fluorescent lighting is used in workshops having rotating machinery, altkrnate fixtures should be connected to different phases to prevent a stroboscopic effect. 5.10 MachinerySpace Lighting. The general illumination in machinery spaces is usually supplied by multilamp fluorescent deck fixtures utilizing 20- or 40-watt lamps controlled at the distribution panel. The location and spacing is generally dictated by interferences with pipes, vents, and other obstructions. Fixtures should be mounted as high as practicable to give good coverage, but they must be accessible for cleaning and replacement of lamps. Where necessary to avoid interference or shadows, the fixtures are stooled down from the deck. The fixture supports must be designed to avoid vibration and must be rigidly braced. The general illumination is supplemented by&--)-fi$t guarded watertight fixtures located as required by the arrangement of the space. This supplementary illumination is particularly necessary in the bilges and for gages, oil sight-flow glasses, and similar fittings. Switchboard illumination must be designed carefully to provide adequate light for the reading of instruments without reflection or glare. A continuous inverted a trough with a line or lines of fluorescent lamps frosted glass cover provides ideal illumination. Auxiliary machinery spaces, such as compartments for the steering gear Or generally are lighted guarded waterti%ht incandescent fixtures' fluorescent fixtures are used if the spaces are heated. Watertight receptacles, or combination switches and receptacles, are provided in main and a~xiliah'mach3inery spaces for the attachment of portable lights and tools. I n the efigine room, receptacles are spaced about 40 to 50 f t apart, and in the smaller spaces at least one double receptacle is installed. 5.1 1 Cargo Hold Lighting. There are two concepts for cargo hold lighting. I n one the illumination is provided by portable "cargo clusters' only; each cluster consists of a fitting accommodating from one to eight incandescent lamps, installed in a plastic or steel bowl reflector, and protected by a wiremesh guard. Receptacles for the attachment of these portable fixtures are installed a t each cargo hatch. The second, and more prevalent, concept is to have permanently installed fixtures for general illumination supplemented by cargo clusters as discussed in the first concept. Where lighting fixtures are installed on the deck over for general illumination, they should be installed between the deck beams or adjacent to the deep beam surrounding the hatches, where they are more readily protected from damage. By providing the fixtureg with resilient mounts, they are further protected from possible damage due to shock from heavy cargo dropped on the

deck above, from shifting cargo caused by heavy seas, and from rough cargo handling. To reduce theft losses, the fixtures in cargo holds and other spaces accessible to casual laborers should be protected by guards requiring special keys or wrenches for removal, and receptacfe caps should be of the hinged type. Watertight q u b l e receptacles, usually two at each hatch, are provided for the connection of portable lights for use in cargo holds and on the deck in the vicinity of cargo hatches.

5.1 2 Weather Deck and Caw0 Handling Lighting*

Fixtures for open decks are required to be of watertight c o n s t ~ ~ t i o nOn . Passenger promenade decks, they are U S U ~ ~ deck ~ Y fixtures ~ n s i s t i n g of a screwed type holophane or a frosted globe and a cast bronze body without a guard- Else\vhere, they may be standard guarded, watertight deck or b u l k h e a d - t ~ ~fixtures. e Exterior lights visible from ahead should be shielded. Lights on the navigating bridge deck and in other areas which cause interference to navigation should be controlled by switches located in the wheelhouse. Portable floodlights, with local control, are provided PO& and starboard to illuminate the gan&cwaYs. They should be mounted on brackets that are so constructed that they may be s w n g outboard and locked in place for overside lighting. -lighting of the oa'g0,-:handling gear and the sumunding deck, watertight floodlights, usu"~ of 309- to 500-watt rating, are permanently mounted high o ~ m ~ O r - E n g p o s t sThe . lights and particularly the mounting supports must be of rugged design to suit the exposed location. The lights are ordinarily not be set and secured in the position adjustable and which directs the beam to best advantage. 0, boat decks, lights are provided to facilitate the loading and lowering of lifeboats. These should be wide-beam 500-watt incandescent watertight floodlights located on the rail or on a vertical pipe stanchion between each pair of adjacent boats. ~h~ mounting arrangement should permit turning the floodlight inboard for deck illumination or outboard and vertically downward to light the water alongside. A quick-release clamp is pmvide&to secure the light in any position. For identification purposes, the owners usually require floodlights for illumination of the smokestack insignia on each side. Vessels on regular passenger runs may have the name boards lighted for identification and for the . convenience of the public&The trend is to the use of electric-discharge mercuryvapor floodlights for all open-deck applications except for lifeboat handling. 5.13 Navigational Space Lighting. When the ship is underway at night, all lights provided for general illumination in the wheelhouse and all exposed lights in the forward part of the vessel are extinguished for better vision of the navigating officers. Any illumination that would be a detriment to navigation must be properly controlled. J

MARINE ENGINEERING

ELECTRIC PLANTS

In the wheelhouse, only lights built into the binnacle, telegraphs, and other necessary instruments remain lighted. The character and intensity of these lights should be such as to permit reading the instrument without interference with outside vision. A shaded light controlled by a switch is provided over the chronometer box t o illuminate the faces of the chronometers when reading is necessary. I n the chart room there should be provided over the chart table one or two bulkhead-mounted adjustable arm type fluorescent desk lights fitted with a switch and a red filter. The arrangement should be such that the lights can be moved in a horizontal plane over the entire table for intense lighting of any portion. The log desk in the wheelhouse should have a similar light. 5.14 Navigation and Signaling Lights. All vessels must be equipped with running and signal lights in compliance with applicable International and Inland Rules of the Road as enacted into law by the United States Congress. The United States Coast Guard publication "CG-169 Rules of the Road, InternationalInland" provides detailed requirements regarding these lights. All fixtures are watertight, substantially constructed of corrosion-resisting material, and are fitted with Fresnel lenses. Navigation and signal lights fitted on ocean-going vessels are listed in Table 3. A typical arrangement of navigation and signal lights for a cargo vessel is shown in Fig. 15. The side lights are located port and starboard on the navigating bridge level so that each is visible at a distance of at least 2 miles on its respective side from

directly ahead to 22.5 deg abaft the beam. These lights are fitted with inboard screens projecting at least 3 f t forward from the lights so as to prevent these lights from being seen across the bow. The masthead and range lights are identical and screened so as to show forward through an arc of 225 deg; that is, to 22.5 deg abaft the beam on either side. These lights are located usually one on the forward mast and one on the after mast, in a line with and over the keel so that the forward light (masthead) is a t least 15 f t lower than the after light (range) and visible at a distance of at least 2 miles. The horizontal distance between the two lights must be at least three times the vertical distance between the lights. For an on-coming vessel the alignment of masthead arid range lights shows the course, and the color of the visible side light indicates whether itwill pass to starboard or port. If both side lights are visible, the vessel is coming head-on. The stern light is located on the centerline at the ship's stern and is screened so as to show aft through an arc of 135 deg (67.5 deg to port and starboard of the centerline); it must be visible from a distance of at least 2 miles. The forward anchor light is located at the bow of the vessel and at least 20 ft above the hull; the after anchor light is located near the stern at a height not l e ~ sthan 15 f t lower than the forward anchor light. Both lights are required to be visible all around the horizon at a distance of at least 3 miles. The forward light is usually permanently mounted on top of the jack staff and the after light on top of the ensign staff. The not-under-command light installation consists of two lights located in a vertical line, one over the other, not less than 6 f t apart and visible all around the horizon at a distance of at least 2 miles. These lights Table 3 Navigation and Signal Lights on Oceangoing Ships are normally portable but may be permanently installed. The towing light installation in accordance with International Rules consists of two lights located in a vertical line with the masthead light, one over the other, not less than 6 f t apart and screened to show forward green 112.5 Starboard side Port aide 112.5 red through an arc of 225 deg (22.5 deg abaft the beam on 225 clear JMasthead either side). On tugboats and other vessels expected to 225 clear Range operate frequently with a tow, thege lights are installed 135 clear ustern 360 clear Forward anchor permanently; otherwise they are portable. After anchor 360 clear The running lights (masthead, range, stern, and side) Wot under command 360 red must be constructed so as to have a backup light in event 225 Towing dear

NAVIGATION SEARCHLIGHT -STERN

LT.

1=.---RANGE

LT. NOT-UNDER-COMMAND

----S I G N A L SEARCHLIGHTS (BRIDGE WINGS)-'

L

~

MASTHEAD LT.

LTS.

LTS.l PORTa ~ ~ STBD.

Fig. 15 Typical arrangement of navigation and dgnal llghh f a a cargo vessel

I

of failure of the primary light. One of the two methods used for complying with this requirement is the use of a bi-filament lamp in a single-compartment fixture. The other method is the use of a two-compartment fixture with a single-filament lamp in each compartment. The latter method is considered the most dependable and is the most commonly used. The signal lights (anchor, not-under-command, and towing) are singlecomparb ment with a single-filament lamp. Each compartment of single and two-compartment fixtures is fitted with a No. 14 AWG &conductor flexible cable having a 3point plug for attachment to its respective lighting supply receptacle. These plugs and receptacles should be of a unique design to preclude their use for other purposes. Vessels equipped with a steam whistle are usually provided with a whistle light to illuminate the steam released when operated. In addition to the foregoing, tank vessels are required to display a red warning light during transfer of bulk cargo; this light should be located above the wheelhouse so as to be visible on all sides. The exact location of running and signal lights and the provision of mounting arrangements, screens, and rigging are generally the responsibility of the hull designer. 5.15 Navigating and Signal Light Controls. A navigating and signal light panel is installed in the wheelhouse for control of all running and signal lights. This panel combines an automatic or semiautomatic telltale navigating light section for audible and visual alarm and' control of the masthead, range, side, and stern lights and a signal light section for control of the anchor, not-under-command, and towing lights. The navigating light section is arranged to indicate failure of each primary lamp or filament and is provided with a switch, either automatic or manual, for transfer to the secondary lamp or filament. The signal light section is arranged only for "on-off" operation of the individual light supply circuit. 5.16 Signaling Lights and Searchlights. A daylight signaling light is required on all ocean and coastwise self-propelled vessels over 150 gross tons and on tankers over 150 gross tons that are engaged on international voyages. It may be either a portable hand-held type, a permanently fixed and wired lZinch unit mounted on top of the wheelhouse, or a semi-fixed 12-inch unit with arrangements for quick mounting and electrical connections at either wing of the navigating bridge. The portable unit may be complete with a self-contained battery or energized from a special low-voltage receptacle.

Searchlights, as such, are not required by the regulatory rules and are installed only when requested by the owners. When requested, u+ly an l&in. incandescent searchlight is installed on the top of the wheelhouse with manual control of train and elevation from within the wheelhouse. ~p"on-offn switch would be located near the operating point. Ships traversing the Suez Canal are required by the Suez Canal Authority to have mounted over the bow a searchlight with a special diffused beam to illuminate the banks of the Canal. Normally, this light is rented from the Canal Authority and is energised from a special receptacle located at the bow. 5.17 Bmcket Fans. Bracket fans have become obsolescent in living areas due tq the extensive use of air conditioning; however, they are used in spaces not air conditioned mch as galleys, pantries, laundries, workshops, and the wheelhouse. Bracket fans are either 12 in. or 16 in. and of the oscillating, 3-speed marine type. Mounting on thin partition bulkheads, where vibration may result in objectionable noise, should be avoided. Receptacles are provided for each fan and are located so that the portable cable length is minimal. 5.1 8 Wiring Appliances. The bodies of watertight fixtures and the special mounting boxes of other fixtures are used for wiring connection boxes as far as circumstances permit. Additional branch boxes are provided where necessary. For exposed wiring, they should be of watertight or nonwatertight type as required, and surface mounted. I n ceiled spaces, they may be flush mounted with sheet metal covers if the decorative character of the space permits; otherwise, they should be installed behind removable panels for accessibility. Switches and receptacles in public spaces, and living areas in general, should be of standard commercial type in the smallest enclosure obtainable; the depth, in particular, is limited. Switch and receptacle plates are usually of a standard sise but may be a special narrow type to fit on joint strips and door frames. They may be of brass or aluminum suitably finished, or plastic material. The mounting boxes for switches should be of adequate size for wiring and are provided with arrangements for clamping the entering cables. Specially designed boxes to suit conditions may be required. In spaces requiring watertight installation, the appliances should be-of the standard marine h t e a i h t tGe. Where aluminum boxes are used, particularly in the weather, care should be taken in their mounting to avoid the possibility of deterioration through electrolysis.

640

ELECrRlC PLANTS

MARINE ENGINEERING

I

Section 6 Lighting and Power Distribution 6.1 General. Energy for lighting and power loads is supplied from the ship service generators (through their associated switchboard and via the ship service distribution system) or from the emergency generator or battery (through the emergency switchboard and via the emergency distribution system). Normally, the emergency switchboard and the emergency distribution system are energized through a bus tie from the ship service switchboard. If the ship service power fails, the emergency distribution system is automatically transferred-fromits normal source to the emergency generator. Subsection 3.7 contains details regarding this automatic transfer. There are many different arrangements for distributing power to the various types of electrical loads installed on shipboard. Figures 16, 17, and 18 are typical generator and bus tie diagrams for a tanker, a dry cargo or containership, and a large passenger vessel, respectively. It should be noted that these diagrams are typical in layout and should not be considered as being restrictive or fixed for any type of vessel.

On large passenger vessels two or three sub-distribution, or "load-center" switchboards are provided for lighting and power distribution. Generally, one will be located in the forward part of the vessel, one aft and, if the size of the vessel warrants, a third will be amidship. Each should be centrally located regarding the loads supplied. Each of the load-center switchboards is supplied from the ship service switchboard by a bus feeder, as shown by Fig. 18. This arrangement is much more economical than providing numerous long feeders from the ship service switchboard to all parts of the vessel. Each load-center switchboard should be installed in a suitable compartment. These compartments usually serve also as centers of electrical serviceand maintenance, and each may be provided with a workbench and with bins 6nd a locker for spare lamps, fuses, and other electrical supplies. 6.2 Lighting Distribution. The lighting bus of each distribution switchboard is supplied by a threephase transformer bank; each bank consists of three 450/12U-

volt singlephase transformers connected delta-delta. In some installations that use rapid-start (without starters) fluorescent lamps, the secondary of the transformer bank is wyeconnected, in lieu of delta-connected, re to with the neutral-_grou ensure reliab6tarting o All lighting distribution panels are supplied by threephase feeders from the lighting bus of the applicable distribution switchboard. These panels are arranged for a three-phase supply and singlephase distribution. The single-phase loads are connected to the threephase aupply bus so as to ensure approximately balanced loading per phase. a. Lighting Feeders. All ship service lighting requirements are supplied by feeders from ship service distribution switchboards through lighting distribution panels. In general, it is economically good practice to limit the load supplied by each lighting feeder to less than 100 amps, so that the feeder may be supplied from its bus through a 100-amp circuit breaker. At least two feeders are provided to senre the lighting requirements of each machinery space.

Separate feders are provided for lighting in cargo spaces. One feeder is usually provided for each cargo hold so they can be disconnected at the switchboard when the vessel is a t sea, thus eliminating the possibility of electrical fire hazards in thpse unmanned spaces. Separate feeders are provided as necessary to supply all lighting requi7ements in working and living areas not covered in the foregoing. For passenger vessels that are subdivided into zones by fire-screen bulkheads, separate feeders are provided for each zone as necessary to supply the lighting requirements between adjacent fire-screen bulkheads. Ship service feeder and emergency feeders that supply the same or adjacent areas sxould be routed so as to be separated as widely as practical to minimize the possibility of damage to both feeders from the same casualty. q e based on 100 percent of Lighting. feeder c-a,b&-..ggs the total connected load plus t6e +errqge-iF,tive circuit load for each spare switch or circuit breaker on the panel being supplied. b. Location of Lighting Panels. For machinery spaces, the ship service lighting panels are usually /

CIRCUITS TO INDIVIDUAL LIGHTING LOAD CENTER FWD CARGO HOLDS

1

3 0 . 4 5 0 / 1 2 0 VOLT TRANSFORMER LOADS AS REQ'D

I

t

1 I

I

I

I

I

I

I

I

I

I

1

I

I

I

4

I

I

--

i

LOADS AS REQ'D

INTERLOCK

TO CARGO H O ~ D LTG PANELS AS REQ'D

I

-

LIGHTING AFT CARGO LOADHOLDS CENTER

-- -

---

j

I I 2 0 VL !T

EMERGENCY BUS

4 5 0 VOLT EMERGENCY BUS

--

--

I

LOADS AS REQ'D

I

+

I

I

L REQID o A m

t

4

L,

icLonos As R E Q Q ~

EMERGENCY SWITCHBOARD

2s

I

! I

!

I I II

90 I

!

_1

I

SHORE POWER TERMINALS

1

!

I

EMERGENCY SWITCHBOARD

3 0 , 4 5 0 1 1 2 0 VOLT TRANSFORMER

II

120 VOLT BUS

/ - r e

I

I

-+-+TO CARGO HOLD LTG PANELS AS REQ'D

I

r

CIRCUITS TO INDIVIDUAL LOADS

I

AS REQ D AND DI$TRIBuTIoN PANELS

I

- - --- I

,---A

-AS REQ'D

I

120 VOLT BUS 120 VOLT MAIN BUS

--

4 5 0 VOLT MAIN BUS

--

4 5 0 VOLT BUS

- ---

. ----

SHIP SERVICE SWITCHBOARD

SHIP SERVICE SWITCHBOARD

Rg. 16 Typical generalor and bur tie diagram f a tankers

230 VOLT BUS'

I

Fig. 17 Typical generator and bus tie diagran f a a dry cargo a containership

J

I

642

I

MARINE ENGINEERING

located on the main operating level, and where readily accessible as typically shown in Fig. 3 of Chapter 18. On some vessels these panels are located at the main entrance to the space. Panels for cargo lighting are usually located in cargo handling machinery deckhouses so as to be accessible, and so that the +igh+ing in each cargo hold may be deenergieed when loading is complete. These panels are not permitted to be located in the cargo holds. The number of panels for cargo lighting depends on the size and arrangement of the vessel; generally, one panel is provided for each cargo hold. The location of ship service lighting panels in paBsenger and crew spaces is determined, to a degree, by the structural and fire zone subdivision of the vessel. Generally, there will be one or more panels on each deck in each subdivision or fire zone; however, two or more decks may be supplied by a single panel, if the arrangement permits. Each panel should be located as near as practicable to the center of the areas served to limit the voltage drop in the branch circuitsj panels are usually installed on passageway bulkheads. In way of joiner work they should be of the flush type. I n public spaces the panels are located near the en-

ELECTRIC PLANTS

trance doors and where the operator can see the lights controlled. c. Lighting Branch Circuits. Lighting branch circuits may be 15-amp, 20-amp, or 30-amp capacities depending on the specific application. Fifteen-amp branch circuits are used for general lighting applications, and each circuit is limited to a maximum connected load of 12 amps (1380 watts) when wired with a No. 12 AWG conductor. When wired w i t h u E G coii&ctors, the maximum connected load is 880 wet&: - --Twenty-amp branch circuits are normally used to supply only fixed nonswitched lighting fixtures for cargo hold or deck lighting. Each circuit should be limited to a maximum connected load of 16 amps and be wired with not less than No. 12 AWG conductors. Thirty-amp branch circuits are normally used to supply only fixed nonswitched lighting fixtures having lamp holders of the mogul (oversize) type or lamps exceeding 300 watts. Each circuit should be limited to a maximum connected load of 24 amps and be wired with not less than No. 10 AWG conductors. Appliance loads, heater loads, and miscellaneous small motors utilizing lighting system voltage may be supplied from lighting distribution panels. Each branch circuit

should be limited to a maximum connected load of 30 amps. The connected load on a general lighting branch circuit is based on the actual lamp (incandescent) sizw installed, but not lei% than 50 watts per lamp unless the fixture design does not permit the use of lamps having a higher wattage than originally installed. The connected load for circuits supplying electric-discharge type lamps (fluorescent and mercury vapor) is based on the ballast i n ~ u tcurrent for each fixture. Receptacle outlets in'stalled for the convenience of the passengers or crew are not included as a connected load. Special lighting fixtureshaving a large number of lowwattage lamps are supplied by a three-phase circuit when the total load of the fixture exceeds 12 amps. The supply circuit is controlled only from the distribution panel, and the current in any conductor is limited to 12 amps. Overcurrent protection for lighting branch circuits is limited to 10-amp fuses or Isamp circuit breakers for the 880-watt circuits, I s a m p fuses or circuit breakers for the 1380-watt circuits, and 20-amp and 30-amp fuses or circuit breakers for the 20-amp and 30-amp circuits respectively. In general, lighting branch circuits in machinery spaces are arranged with alternate groups of lights on different branch circuits so that large areas wiU not be B'A 1- EY ' /3%450/120 VOLT r-7~ CHARGER 1 i put in darkness by failure of a single branch circuit. In I r TRANSFORMER these spaces no individual switches are provided as the SWITCH~OARO . FWD SHIP SERVICE SWITCHBOARD lights are controlled only by the panel switches. 4 51 0-0" (20;;BUS 4 5 0 VOLT BUS " 7 Each passenger stateroom and public space should be supplied by at least two ship service lighting branch circuits, so arranged that in event of failure of one branch there will be adequate light to permit use of the space. The ceiling fixtures in each stateroom, berthing space, storeroom, and similar small compartments are controlled by a switch at the room entrance, and located about 4 ft I .I . L I I i ~ i ~ i above the deck. Staterooms having more than one I I -I I entrance door should have a switch at each door for t t LOADS AS control of the ceiling fixtures. Berth, desk, lavatory, IC LOADS LOADS AS LOADS AS 1 i?) I REO'D AS REO'D' REO'D REO'D and dreeaing table -lights are usually controlled by I I individual switches mounted on the fixtures or located near the fixture. r--/30.450/120 VOLT, XFMR, ! I Separate branch circuits are provided exclusively for passageway lighting. Also, the lights in each passage should be divided between ship service branch circuits and emergency branch circuits, such that normal and emergency illumination requirements are satisfied. Separate branch circuits are generally provided for bracket fans; however, where relatively few widely distributed fans are installed, they are supplied from local lighting branch circuits. Branch circuits should not be routed through firescreen or watertight bulkheads. -. Typical arrangements of lighting branch circuits for L. . . . I I . . .. L.I . .1 I . ..1 . AFT LQhD CENTER SWBD a crew stateroom and a passenger stateroom are shown AFT SHIP SERVICE SWITCHBOARD FWD LOAD CENTER SWBD MIDSHIP LOAD CENTER SWBO in Figs. 19 and 20 respectively. Fig. 18 Typical wmratw and bur tie diagram for Q large panengar v-l 6.3 Power Dirhibvtion. Ship servicQpower system

-

1 '

loads consist principally of motor-driven auxiliaries and heating equipment, and are supplied either individually or in groups by feeders from a ship service distribution switchboard. Individud feeders are normally used to energize large propulsion pbnt auxiliaries that are located in the same space as the distribution switchboard, but may ,be used for large motors anywhere in the vessel. Grouped loads are supplied by feeders through distribution panels, these panels being centrally located to the loads supplied. a. Power Feeders. Separate feeders should be provided to panels and group control boards serving machinery space auxiliaries and refrigeration equipment that are not supplied individually. Machinery space ventilation fans, living and working space ventilation fans, and cargo space ventilation fans should be supplied by separate feeders. Each ventilation feeder circuit breaker should be provided with a means of remote control for deenergizing its feeder in case of fire; the remote control device provided for deenergizing msr chinery space ventilation feeders should be located in the passageway leading to, but outside of, the machinery space. For all other ventilation feeders, the remote control devices are usually located in the wheelhouse, but may be located in a space near the wheelhouse as permitted by the regulatory rules. The means of remote control for ventilation feeders,$onsists of a normally closed switch which, when 'operated to the "stop" position, deenergbes an undervoltage trip device on the circuit breaker, thus tripping the circuit breaker. This arrangement may be such as to trip several breakers (feeding various ventilation feeders) at one common trip switch. These switches are mounted in special locked enclosures having glass fronts which must be broken to operate the switch. Separate feeders should be provided for galley appliances, air heaters other than isolated units, and each group of cargo handling equipment. Equipment required to be operated when underway should not be supplied from the cargo handling equipment feeders, since these feeders are usually disconnected from the Wribution switchboard when at sea. Windlass and capstan motors may be supplied from these feeders if convenient. The steering gear should be supplied by two independent feeders, separated as widely as practical to reduce the probability of 10- of power from a single casualty. Both feeders are pormally applied from the ship service distribution switchboard; however, if the capacity of the emergency power source is adequate, one feeder may be supplied from the emergency switchboard. The steering gear distribution panel is arranged, through interlocked circuit breakers, to connect each motor or appliance to either feeder. A separate feeder should be provided from the ship service distribution switchboard to a shore connection box. The shore connection box should be suitably located for supplying power from shore to ship when tied up at shore facilities.

'

644

ELECTRIC PLANTS

MARINE ENGINEERING

BERTH LTS. 30 W EACH FAN 75W

11

LEGEND 0 RECEPTABLE X SWITCH OQ DOUBLE RECEPTABLE K SWITCH, 3 WAY Fig. 19 Typical lighting branch circuit for a crew stateroom

O< -SWITCH, 3 WAY @ -RECEPTACLE

8 -CONNECTION BOX

Fig. 20 Typical lighting branch circuit for a pauemger stateroom

The minimum current-cawing capacity of power system feeders to individual motors should be based on a t least 125 percent of the motor full-load current. In general, for feeders supplying a group of motors, the feeder current capacity should be based on 125 percent of the largest motor rating, plus the sum of the ratings of all other motors, plus 50 percent of the rating of the spare switches on the panel supplied. For feeders supplying cargo handling equipment, workshop tools,

windlass, and capstans, refer to regulatory rules for specific requirements in determining cable currentcarrying capacities. The current capacity of galley panel feeders should be based on 100 percent of the first 50 kw of load or one half of the total connected load (whichever is larger), plus 65 percent of the remaining connected load, plus 50 percent of the ratings of the spare switches. The current capacity of transformer feeders should be based on

100 percent of the rated primary and secondary currents. Steering gear feeder current capacities should be based on the total ratings of all equipment normally connected to it that operate simultaneously. b. Location of Power Panels. For mchiinery spaces, the ship service power panels, including those for ventilation fans, are usually located on the main operating level, so as to be readily accessible and central to the auxiliaries supplied. Panels supplying cargo handling equipment and cargo hold ventilation panels are usually located in cargo handling machinery deckhouses. The steering gear transfer panel is located in the steering gear - room. Ventilation panels serving living and working spaces, and general service panels are located as near as practical to the loads supplied. Where the loads served are located in a single compartment, the panel should be in the same space. Galley panels are located within the galley space or in a passageway and adjacent to the galley entrance. c. Power Branch Circuits. Separate branch circuits should be provided for each motor having a full-load current of 6 amps or more, and for each air heater regardless of rating. With these exceptions, several small loads, having a total rating not exceeding 7.5 amps, may be grouped and supplied by a branch circuit through a subdistribution panel; the circuit conductors should not be smaller than No. 14 AWG and should be protected by a circuit breaker of not more than 15-amp rating or fuses or not more than 10-amp rating. Also, groups of loads not exceeding 15 and 20 amps respectively may be supplied by branch circuits as noted in the foregoing, provided the circuit conductors are not smaller than No. 12 and No. 10 AWG, and the overcurrent protection is rated not over 15 and 20 amps respectively. Receptacle outlets should not be supplied from these branch circuits. The current-carrying capacities of motor .branch circuits should be based on at least 125 percent of the motor full load; the carrying capacities of all other branch circuits should b e based on a t least 100 percent of the connected full load. All branch circuits should be wired with not less than No. 14 AWG conductors. Each branch circuit should be protected by a circuit breaker, with thermal or magnetic trips, or by a fuse. Since there are many different requirements applicable to overcurrent protection for branch circuits, the regulatory rules should be reviewed for each specific application., The magnetic instantaneous setting for motor branch circuits .should be a t a higher value than the starting current of the motor; for branch circuits supplying loads through transformers, this setting should be higher than the inrush current of the transformer. Power branch circuit loads that require other than normal ship service voltages are supplied either by individual transformers for each load, or by a bank of transformers in the feeder circuit just ahead of the distribution panel. Typical loads of this type include

645

laundry equipment, galley equipment, and special refrigerated containers. . d. Special Features. Vent fans serving spaces that are subject to carbon dioxide flooding should be arranged to be deenergized when carbon dioxide is released into the space as the carbon dioxide will otherwise be spread to other regions of the ship. The means for deenergizing the fan motors is a 'presstireoperated switch in the carbon dioxide piping system. When the switch is actuated, by relgase of the carbon dioxide, it may deenergize an uidervoltage trip device on the circuit breakers supplying the fans involved, thus tripping the circuit breaker and stopping the fan motors, or actuation of the switch mav cause the fan motor control circuit to be deenergized, thus stopping the fan motor. Also, a similar emergency means of stopping fuel-oil service pumps, fuel-oil transfer pumps, and forced-draft blowers should be provided if the machinery space in which they are located is subject to carbon dioxide flooding. Stopping these motor-driven auxiliaries would avoid the possibility of fuel oil feeding the fire from a ruptured oil line, and reduce the probability of the forced-draft blower discharging the carbon dioxide from the space. Circuit breakers supplying the fuel-oil service pump, fuel-oil transfer pump, and forced-draft blower motors should be provided with a means of remote control for deenergising each motor in case of a fire in the machinery space. This means of remote coiitrol may consist of a normally closed switch which, when operated to the "stopJ' position, deenergizes an undervoltage trip device on the circuit breaker, thus tri$ping the circuit breaker. This arrangement may be such as to trip each breaker ~ These switches are mounted at one common t r i switch. in special locked enclosures having glass fronts which must be broken to operate them. The switches are required to be located outside the machinery space and may be located adjacent to the emergency stop switches provided for machinery space ventilation fans. The Coast Guard rules should be reviewed to determine the specific requirements, as may be applicable to each vessel, regarding the following: watertight door system, fire-screen door system, lifeboat winches, steering gear, air heaters, motion picture projection rooms, hospital operating room, and locations where gasoline or motor fuel is carried in vehicles. These requirements involve: types and location of equipments, special codes that must be complied with, cable size and overcurrent protection restrictions, and special features for control circuits. For a discussion regarding the automation of the machinery space, includins bridge control of the propulsion plant, see Chapter 51. On many vessels, an impressed current cathodic protection system is provided to prevent corrosion of the rudder, propeller, and the submerged hull. The system supplies a low positive voltage to submerged anodes from which current flows through the seawater to the hull, the purpose being to suppress the normal flow of small currents between areas of a submerged

646

ELECTRIC PLANTS

MARINE ENGINEERING

hull so as to prevent corrosion. System components usually include : 1 Hull-mounted reference electrodes, in quantity as required, to provide the degree of protection desired. 2 Hull-mounted anodes, in quantity and location as required, to provide the desired protection. 3 Controller, usually located on the operating level in the main machinery space, to select, monitor, and provide the proper signals to the power supply units for the desired protection. 4 Power supply units, in quantity and location as necessary, to provide the necessary regulated d-c power to the anodes. 5 Propeller shaft grounding assembly to provide a path of low electrical resistance from the propeller to the hull. 6.4 Emergency Lighting and Power. Some form of emergency lighting ordinarily is provided for every vessel equipped with an electric lighting system. Exceptions include, (a) small passenger vessels that operate only between sunrise and sunset, and (b) small passenger vessels that operate not more than ore of having their source of general lighti the propulsion plant and located above the bulkhead deck. The temporary source of emergency power required on large passenger vessels is of limited capacity and is rated to supply only selected emergency loads, for a short time, while a large-capacity final emergency source is being started. J X e following loads are reqhired to be energized from the tgmporary source when available: 1 Navigation lights. 2 Adequate number of lights in the machinery space to permit the performance of essential operations and observations under emergency conditions and to facilitate the restoration of service. 3 Lighting for passageways, stairways, escape trunks, passenger quarters, crew quarters, public spaces, machinery spaces, and work spaces adequate to permit the passengers and crew readily to find their way to open decks and to lifeboat embarkation and assembly points with all watertight doors and fire-screen doors closed. 4 Illuminated signs bearing the word "exit" in red letters installed as required so that the direction of escape to the open deck is apparent (passenger vessels only). 5 General illumination for safe operation of poweroperated watertight doors. 6 One or more lights in galleys, pantries, radio room, steering gear room, emergency power room, chart room, wheelhouse, crew's mess, and recreation rooms. 7 Lighting for boat and embarkation decks and passenger assembly points for safe embarkation into lifeboats. 8 Electric communication systems essential under temporary emergency conditions and which do not have an independent storage battery source of power. 9 Power-operated watertight door system. 10 Emergency loudspeaker system.

11 Fire-screen door holding and release system. 12 Supply to motor generator or other conversion equipment when a temporary source of a-c power is necessary for essential communication systems, emergency, or safety requirements. The entire emergency load, consisting of the following, is required to be supplied from the final source of emergency power : 1 All loads energized from the temporary source. 2 Illumination for the safe operation of the lifeboat and life raft launching gear and the lifeboats and life rafts in process of and immediately after launching. 3 Charging panel of temporary emergency battery and of starting battery for the diesel engine driving the emergency generator. 4 One bilge pump, one fire pump, and one sprinkler pump, if required to be supplied from the emergency source. 5 Daylight signaling lights. 6 Smoke detector system. I n addition to the aforementioned requirements, regarding the loads supplied from the final source of emergency power, the United States Coast Guard rules recommend that the following loads, if installed, be supplied from the normal emergency source when the capacity and character of the emergency source permits:

Radio installation Radio direction finder Loran ,+A Radar plan position indicator \g Gyrocompass Depth sounder 9 Electric whistle and siren control The emergency source of power is required to have sufficient capacity to supply only those loads that are required to have an emergency supply; however, nonemergency loads may be supplied from the emergency source provided there is adequate capacity to supply all of the loads that may be connected simultaneously. The location of the emergency generating set and its characteristics are described in Subsections 2.4 and 2.5, respectively. When the emergency power source is a battery, its location requirements are the same as noted for the emergency generator. Batteries used for the emergency power source should have characteristics as specified in the rules of the regulatory bodies. The arrangement of the emergency switchboard, including automatic transfer equipment, is discussed in Section 3.7b. The requirements for emergency supply for various types of vessels are included in the United States Coast Guard rules. The Coast Guard rules are revised periodically, but are typically as follows:

a. Passenger Vessels Over 65 ft in Length 1 I n ocean and coastwise service, the emergency source should be a storage battery of sufficient capacity to supply continuously the full emergency load for 36 hr, or

a diesel or gas turbine driven emergency generator with capacity to supply continuously the full emergency load for 36 hr and a temporary source of emergency power consisting of a storage battery of sufficient capacity to supply the "tem~orary"emergency load for not less than hr. Arrangements must be provided for an automatic transferof the full emergency load to a battery supply upon loss of the normal supply, or an automatic transfer of the U t e m ~ o r a load ~'l the battery s uand ~ ~ automatic starting of the emergency generator, transfer of the full emergency load to the emergency generator upon loss of the normal supply, as applicable. 2 In than Ocean and coastwise service, for vessels of 100 gross tons and Over, the source should be a diesel or a gas turbine driven emergency generator or a storage battery of sufficient capacity supply ~ontinuousl~ the full emergency load for 8 hr, or twice the time of run, whichever is the smaller. Arrangements must be provided automatic transfer of the load to battery upon loss of supply, or automatic shrting ofthe emergency generator and the transfer of the emergency load to the emergency generator upon loss of normal supply, as applicable. 3 In other than ocean and coastwise service, for vessels over 15 gross tons but less than 100 gross tons, the emergency source should be a diesel or a gas turbine driven emergency generator of a storage battery of sufficient capacity to supply continuously the full emergency load for 8 hr or twice the time of run, whichever is smaller. Arrangements must be provided for either automatic (as just noted in a.2) or manual transfer of the emergency load to the emergency source. Manual transfer requires that only a single operation of an "onoff" switch will cause the emergency system to supply its connected load. This "on-off" switch is located in the wheelhouse or as necessary to be under the control of the chief engineer.

b. and and Tank Ships

Se&?opeued

1 In all-waters service, for vessels of 1600 gross tons and over the emergency source is the same as just given in a.3 except that the emergency load is required to be camed for 12 hr. 2 I n all-waters service, for vessels of 300 gross tons and over, but less than 1600 gross tons, the emergency source may be the same type as noted in a.3 or may be relay-controlled battery-operated lanterns. The emergency source selected must have sufficient capacity to supply the full emergency load for 12 hr or twice the time of run, whichever is smaller. When batteryoperated lanterns are used for the aforementioned emergency service, they are required to have rechargeable batteries, incorporate an automitic battery charger that w i l l maintain the battery in a fully charged condition, and not be readily portable. Also, the minimum period of operation of these lanterns may be less than 12 hr but not less than 6 hr.

6.5

647 Emergency Distribution System

a. Lighting. In general, the emergency lighting system forms a part of the ship servicelighting system and is energized at all times when the passengers or crew are aboard. Separate emergency lighting feeders are provided for machinery spaces, crew and passenger areas, and lights controlled from the wheelhouse. The machinery space feeder supplies only those emerlgency ~ , lights loc,ted in the machinery space; this rupply is through a didribution panel usually located above the operating l e e 1 and near the main entrance to the space. Usually the crew and passenger area feeders supply distribution panels suitably located for the control and distribution of the emergency fights throughout these areas. For passenger vessels subdivided into mnes by firescreen bulkheads, a t least one emergency feeder is provided to supply the lighting in each zone between adjacent fire-screen bulkheads. A separate feeder, supplying a panel in the wheelhouse, is provided for emergency lights located in or controlled from the Loads supplied from this pane] include the navigation light panel, signaling lights, emergency lights on open decks, lifeboat lights, and chart room lights. The navigation light panel is generally supplied by a through feed, without protection, from the wheelhouse emergency panel; but, as an alternative, the navigation light panel, signaling lights, and lifeboat lights may be supplied by separate feeders from the emergency switchboard. I n addition, on vessels that have two "islands" (e.g., those with midship and aft houses), lighting for lifeboat-associated areas located remotely from the wheelhouse "island" may be controlled from a central location within the remote "island " in lieu of from the wheelhouse. Floodlights for adjacent lifeboats should be supplied by different branch circuits. Emergency lighting fixtures usually are of the same type as the ship service lighting fixtures installed in the same spaces. I n public spaces where fixtures may contain more than one lamp, only the necessary number of lamps to give the required illumination are connected to the emergency circuit. For ready identification, all lighting fixtures on the emergency system are identified by a metal tag stamped with the letter E (in red) a t least % in. in height. This tag is secured to the deck or paneling immediately adjacent to the fixture and, for fixtures having lamps on both the ship service and emergency circuits, an additional tag is secured to the emergency supply cable within the fixture mounting box. b. Power. Separate &ergency power feeders are routed from the emergency switchboard to each power system load individually. Only a limited number of power system loads are required to have an emergency supply and they are not centrally located with respect to each other. c. Communication and Alarm. It is general practice to supply vital machinery space indicating, monitoring,

648

Table 4

Correction Factor for Cable Calculations

CABLE AWG

POWER FACTOR OF LOAD

fl

No.

1.00

0.95

0.90

0.85

9.80

0.75

0.70

0.65

2 1 n

1.00 1 .OO

1.01 1.03 1 0.5

0.99 1.01 1.04

0.96 0.98 1.02

0.92 0.95 0 .gg

0.84 0.88 0.93

0.76 0.80 03 5

0.68 0.71 0.77

r nn

ELECTRIC PLANTS

MARINE ENGINEERING

,

The percent voltage drop in 3-conductor, 3-phase circuits 52,600 CM and smaller is 173 I R L Voltage drop = (2) CMV

92.2' TO CALCULATED -----LOAD CENTER . 8 9 0 % DROP--

n

---

For 3-conduchr, %phase circuits 661400 CM and larger the percent voltage drop is 173 I R L CF Voltage drop = (3) CM V

L and alarm loads from an emergency panel that is conveniently located on the machinery space operating level. This panel is supplied directly by a feeder from the emergency switchboard. Other communication and alarm loads are supplied individually from the emergency switchboard. d. Electronics. Electronic loads are usually energized from an emergency panel, centrally located to these loads, that is supplied by a feeder from the emergency switchboard. Electronic loads include radio, radio telephone, radio direction finder, loran, and depthsounding equipment. Radar equipment is usually supplied directly from the emergency switchboard, since this equipment often requires a power supply voltage different from the remaining electronic loads. e. General. The distribution panels used for emergency system distribution are of the same type as those for the ship service distribution system. Emergency system cables that are not required to terminate at equipment located within machinery spaces, uptakes, or casings should be routed to avoid penetrating the boundaries of these spaces, and should be kept clear of the decks and bulkheads forming these boundaries. Voltage drops in emergency circuits are calculated in the same manner as for ship service circuits. For lighting circuits the voltage drop from the emergency switchboard lighting bus to the most remote fixture should not exceed 3 percent. For all other circuits the total voltage drop should not exceed 5 percent when supplied from the main bus of the ship service switchboard through the bus tie to the emergency switchboard.

CM V CF

m

--

2.17AMPS .#v

-"

.483%DROP

1-

m

-".178%

.--

-

".104%

-1-

-- -

.077%

'

.G8%

1

k---FIVE

5 0 WATT 115 VOLT LAMPS-CALCULATED VOLTAGE DROP . 8 9 0 % ' F 0 ~ LONGEST BRANCH CKTNOTE: 4 4 9 7 CM CABLE USED IN