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Heating, Ventilating, and . Air Conditioning ANALYSIS and DESIGN Fourth Edition

1 'I

Faye C. McQuiston

-J

OklaJwma State University

Jerald D. Parker

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.

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OklaJwma Christian University of Science and Arts

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~ John Wiley & Sons, Inc.

New York" Chichester 0 Brisbane" Toronto 0 Singapore

ACQUIsmONS EDITOR MARKE:riNG MANAGER SENIOR PRODUcrrON EDITOR MANUFACfURING MANAGER lLLUSTRAnON COORDINATOR

Cliff Robichaud Susan Elbe Ingrao Associates and Nancy Prinz Andrea Price Jamie Perea

This book was set in Times Roman by GeneraJ Graphic Services and printed and bound by Hamilton Printing Company. The cover was printed by New EnglaDd Book Components, Inc. Recognizing the importance of preserving what has been written, it is a policy of John Wiley & Sons, Inc. to have books of enduring value published in the United States printed on acid-free paper, and we exert our best efforts to that end.

Copyrigbt'@ 1977. 1982,1988. 1994. by John Wiley & Sons. Inc.

All rights reserved. Published simultaneously in Canada. Reproduction or IrallSlation of any part of this work beyond lhat permitted by Sections 107 and 108 of the 1976 United Slates Copyright Act without the permission of the copyright owner is unlawful. Requests for permission or further infonnation should be addressed to the Permissions Department, Jobo Wiley & Sons. Inc.

I.iJmu;y of Congress CaJaWging.in-hblica&m Data McQniston. Faye C. Heating. ventilating, and air conditioning: analysis and design! Faye C. McQuiston. Jerald D. Parker-4th ed. p. em.

Includes index.

ISBN lJ.471-58 107-0

I. Heating. 2. Ventilation, 3. Air conditioning. I. Parker. Jerald D.

TH7222.M38

1994

697-dc20

Printed in the United 'States of America

109876543

93-2.8394

CIP

Preface

Advances in the areas of load calculations, indoor air quality (lAQ), and the require­ ments for environmentally acceptable refrigerants have prompted this revision of the third edition. The revisions reflect primarily the result of research sponsored by the American Society of Heating, Refrigerating and Air-Conditioning Engineers (ASHRAE) and the continued development of ASHRAE standards related to lAQ. comfort, and refrigerants. The original objective of this book, to produce.an up-to-date, convenient classroom teaching aid based on ASHRAE literature has not changed. It is intended for use by engineering students at the undergraduate and graduate level as well as practicing engineers. Mastery of the material should enable a person to effectively participate in the design of all types of HVAC systems. Basic courses in thermodynamics, heat transfer, fluid mechanics, and dynamics are desirable prerequisites. There is sufficient material for two-semester length courses with considerable latitude in course makeup. Although the book is intended to be primarily a teaching device, it should ,also be useful as a reference and as an aid in studying new procedures. A number of revisions have been made based on suggestions from users of the previous editions. Data and references have been updated throughout the text; however. in a few instances. useful material from older sources has been retained. New problems have been added, existing problems have been revised, and the problems have been rearranged to fit the new order of the material in the chapters. In all major areas there are problems that can be solved either manually or using available computer software. A solution manual is available from the publisher. Instructors should provide some examples and problems that emphasize their own design philosophies and the require­ ments of local geographical regions. Chapter 2, Air-Conditioning Systems, has been revised in an effqrt to improve understanding for the beginning student. Increased emphasis has been placed on the control of HVAC sy'stems by adding ,more explanation of control theory and more detailed illustrations of .typical systems and controls. Chapter 4, Indoor Air Quality-Comfort and Health, has been completely revised to place greater emphasis on i.ndoor air quality and health and the use of ANSUASHRAE Standard 62. The sections dealing with comfort have also been updated to agree with the latest comfort standard,ASHRAE Standard 55. Chapter 8, The Cooling Lo!\d, has undergone extensive revision base,d on the latest ASHRAE research in this area. The Transfer Function Method is more fully explained with sufficient data and examples to enhance understanding of this method. The new CLTD/CLF/SCL manual calculation method is also presented with adequate data and examples. Recent literature and software for load calculations is referenced. All the material related to energy calculations has been collected together in Chapter

9. Chapter 12, Fans and Building Air Distribution, has been revised to reference the new

"

vi

Preface ASHRAE duct fitting database with a sampling of those data. Discussion of evolving duct design optimizing procedures is now included. Mass transfer and direct contact heat transfer material has been condensed and placed in Chapter 13. Chapter 15, Refrigeration. has been revised to include discussion related to ozone depletion. the safety and environmental effects of refrigerants. and the selection of replacement refrigerants. Instructors using this text are encouraged to involve students in the use of personal computers and the many programs available for use. The authors may be contacted for information related to procurement of software. Uncertainty exists as to when a complete conversion from English to the international system of units (SI) will occur in the United States. However, engineers should be comfortable with both systems of units when they enter practice. Therefore, this book: continues to use a dual system of units, with some emphasis placed on the English system. Instructors should blend the two systems of units as they see fit. We are deeply indebted to ASHRAE for providing much SllPport in the production of this book. Many companies and individuals, too numerous to list, contributed sugges­ tions, ideas, photographs, and commentary. Thank you every one. FAYEC. McQmsTON D. PARKER

JERALD

'.

About the Authors

Faye C. McQuiston is Professor Emeritus of Mechanial and Aerospace Engineering at Oklahoma State University, Stillwater. Oklahoma. He received B.S. and M.S. degrees in mechanical engineering from Oklahoma State University in 1958 and 1959 and a Ph.D. in mechanical engineering from Purdue University in 1970. Dr. McQuiston joined the Oklahoma State faculty in 1962 after a three-year period in industry. He was Ii National Science Foundation Faculty Fellow from 1967 to 1969. Dr. McQuiston is 311! active member of the American Society of Heating, Refrigerating and Air-Conditioning Engineers (ASHRAE). recently completing a term as a Vice-President. He has served on the Board of Directors, the Technology, Education, Member, and Publishing councils, is a past member of the Research and Technical, Education, and Standards Committees. He was honored with the Best Paper Award in 1979, the Region vm: Award of Merit in 1981. the Distinguished Service Award in 1984, and the E. K. Campbell Award in 1986. He was also ,elected to the grade of Fellow in 1986. Dr. McQuiston is a registered professional engineer and a consultant to several system design and equipment manufacturing firms. He is active in research related to the design of heating and air-conditioning systems, particularly in the areas of heat-exchanger design and simulation and load calculations. He has written extensively in the area of heating and air conditioning and is the coauthor of a basic fluid mechanics and heat transfer text.

Jerald D. Parker is' a Professor of Mechanical Engineering at Oklahoma Christian University of Science and Arts after serving 33 years on the Mechanical Engineering faculty at Oklahoma State University. He received B.S. and M.S. degrees in mechanical engineering from Oklahoma State University 111 1111100"0

o!)D"OQ.~O ~~O',(!,9 0

0 ¢

»1)'"

x

return e..

(?

Chilled water pump

~ Chilled

;'handlers

water supply

Figure2-1 Schematic of a typical commercial air-conditioning system.

refrigerant. However, a liquid is generally used in commerical applications. Exceptions are discussed later. As, shown in Fig. 2-1, the liquid is usually· cooled with devices referred to as chillers. The chillers are refrigerating machines operating on the vapor compression or absorption cycles (Chapter 15). Heat is rejected to the atmosphere by use of a cooling tower, or the chiller may have an air-cooled c6iidenser as shown. Pumps are required to circulate the liquid through piping, and the liquid cooling equipment is remote from the air-conditioning equipment. To provide heat, a heating fluid must be supplied to the heating coil in the air handler. The fluid is usually hot water, but steam can also be used. Hot water or steam is provided by a boiler at some remote location and circulated by pump and piping. Water may be heated using steam with a heat exchanger, called a converter. The fuel for the boilers may be natural gas, liquified petroleum gas (LPG), or fuel oil. The humidifier is supplied with water vapor or an atomized water spray. Water vapor is most desirable and can be supplied by a steam boiler or a small special steam­ generating· device. '. The components of a complete system will be further discussed below.

16 Chapter 2 2~2

Air-Conditioning Systems

THE AIR~CONDITIONING AND DISTRIBUTION SYSTEM Important first choices in HVAC central system design involve determination of the individual zones to be conditioned and the type and location of the HVAC equipment. Normally, the equipment is located outside the conditioned area in a basement, on the roof, or in a service area at the core of a commercial building. A zone is a conditioned space under the control of a single thermostat. The thermostat is a control device that senses the space temperature and sends a correcting signal if that temperature is not within some desired range. In some special cases the zone humidity may also be controlled by a humidistat. It is most important that the temperature within the area conditioned by a central system be uniform if a single-zone, constant-air­ volume duct system is to be used because air temperature is sensed only at that single location where the thermostat is located. Because conditions vary some in most typical zones, it is important that the thermostat be located carefully, in a spot free from local disturbances and where the temperature is likely to be most nearly the average of the occupied space. Uniform loads are generally experienced in spaces with relatively large open areas and small external loads such as theaters. auditoriums. department stores, and public areas of most buildings. In large commerical buildings the interior zones are usually fairly uniform if provisions are made to take care of local heat sources such as large equipme'ht or computers. Variations of temperature within a zone can be reduced by adjusting local air flows or changing supply air temperatures. Spaces with stringent requirements for cleanliness, humidity, temperature control, and/or air distribution are usually isolated as separate zones within the larger building and served with separate systems and furnished with precision controls. For applications requiring close aseptic or contamination control of the environment, such as surgical operating rooms. all-air~type systems generally are used to provide adequate dilution of the controlled space. These applications usually involve careful control of the diluting air through the controlled,environment space. ,In spaces such as large office buildings. factories. and large department stores, practical considerations require not only multiple zones but also multiple installation of central systems. In the case of tall buildings, each central system may' serve several floors. Large installations such as college campuses, military bases, and research facilities may best be served by a central station or central plants. where chillers and boilers provide chilled water and hot water or steam through a piping system to the entire facility, often through underground piping. For these large installations central plants provide higher diversitY, greater efficiency, less maintenance cost, and lower labor costs compared to individual centrlil facilities in each building. The choices described above are usually controlled by economic factors, involving a trade~off between first costs and operating costs for the installation. As the distance over which energy must be transported is increased, the cost of moving that energy tends to become more significant when compared to the costs of operating the chillers and boilers. As a general rule the smaller systems tend to be the most economical if they move the energy as directly as possible. For example, in a smalI heating system the air would most likely be heated directly in a furnace and transported through ducts to the controlled space. Likewise in the smaller units the refrigerating system wouid likely involve a direct exchange between the refrigerant and the supply air (a D-X system). In

2-3

Central Mechanical Equipment

17

installations were the energy must be moved over greater distances, a liquid (or steam) transport system would likely be used, but this involves extra heat exchange steps. Most commercial systems, such as the one shown in Fig. 2-1, utilize a dual or combination method of transporting energy. In such systems energy is carried from the boiler (or converter) to the air handler heating coil by a liquid, usually water. The energy is then carried to the conditioned space by ducted air. In cooling, energy is carried by return air from the conditioned space to the air handler cooling coil and then transported to the chiller evaporator by a liquid. This energy is rejected by the refrigerator condenser to the ambient air. Where a cooling tower is utilized, energy is carried from the chiller condenser to the cooling tower by a liquid.

2-3

CENTRAL MECHANICAL EQUIPMENT Once the user's needs have been appraised, zones defined, and loads and air require­ ments calculated and the type of overall system has been determined, the designer can start the process of selection and arrangement of the various system components. It is important that the equipment be suited for the particular application, sized properly, accessible for easy maintenance, and no more complex in arrangement and control than is necessary to produce the conditions required to meet the design criteria. The economic trade-off between initial investment and operating costs must always be kept in mind. Consideration of the type of fuel or energy source must be made at the same time as the selection of the energy-consuming equipment to assure the least lifecyc1e cost for the owner. Chapter 15 of the ASHRAE Handbook (2) gives the types and properties of fuels and energy sources and guidance in their proper use. This selection is not only important from an economic standpoint but is also important in making the best use of our natural resources. Ductwork and piping make up a significant part of the design of an HV AC system. These will be described in detail in later chapters. The remaining components can be generally grouped into five categories: ' Air handlers and fans Heating sources Refrigeration Pumps Controls arid instrumentation Familarity with some of t~e components of HVAC systems will make the design and analysis material that follows more meaningful as well as more interesting. A quick look at some of the components in the categories above will now be given. This will be followed by descriptions of some common arrangeinents of these components in modern HVAC,systems.

AirQHandling Equipment

I

i \

The general arrangement of a commercial central system for year-round air-condition­ ing was given iri Fig. 2-1. Most' of the components are available in subassembled sections ready for bolting together in the field or completely assembled by the manufacturer. In the upper-right-hand corner of Fig. 2-1 can be seen the simplified

1

2-3

19

Central Mechanical Equipment

--­

.­

Exhaust

Return

air

air

Damper motors and

power supply

0::

3::

:c

~

-­ -

Outdoor air

IcgJ "

en 3::

:I:

~ ~~-~:

If( .

~ _~Id

~\

air

.!

I

"1

Filter supply fan

tt:

::c

u

cn ::c

u

Individual

zone dampers

I

Figure 2-3 Schematic of a blow-through air handler with hot and cold decks and zone dampers.

1.

cooling coils may be placed in a side-by-side or parallel arrangt;)ment. A schematic of such an arrangement is shown in Fig. 2-3~ The heating and cooling coils are referred to as the hot deck and the cold deck, respecirvely. The arrangement with the fan upstream of the coils is called a blow-through configuration. Figure 2-4 shows a photograpb of a multizone, blow-through configuration. The discharge area of the air handler may be divided so that several zones may be served With separate temperature control in each zone, or the air handler may be used without the dampers in a dual-duct system. BOth of these arrangements will be discussed later in this chapter. Internal components of the central air handler include the cooling, heating, and, for some units, t!W'-.preheat coils. These are usually of the finned-tube type such as are shown in Fig. 14~1O. Coil design and selection will be considered in Chapters 3 and 14. The humidifier ina commercial air handler is usually a type utilizing steam such as is shown in Fig. 2-5. ~e fans are usually centrifugal types, shown in Fig. 2-6. These and other fan types will be looked at in more detail in Chapter 12. A unit-type filter is shown in Fig. 2-7; the various filter types and design procedures will be considered in Chapter 4. Dampers, which can be seen in Fig. 2-4, will be discussed in Chapter 12. The ductwork to deliver air is usually a unique design to fit a particular building. The air ducts should deliver conditioned air to an area as quietly and economically as possible. In some installations the air delivery system consumes a significant part of the total energy, making good duct design and fan selection a very important part of the engineering process. Design of the duct system must be coordinated with the building design to avoid last-minute changes. Chapter 12 explains this part of the system design.

~

II 1

I 1 ~ l

i 1

,!,

1 J

l t

j

1

........... ~ ......

It is clear from Eq. 3-21d that WI is a function of I.,

0.6219~

W;2

J

(3·21b)

* superscript Tefers to the adiabatic saturation temperature, and

= cpa(t; - I,)

I

(3-21a)

(3-21d)

t;, (3-14b)

(P2 - P,,2)

, andpv2 = Ps2 att;· The enthalpy of vaporization IIg2 depends only on t;; the enthalpy of the vapor i vi is a function of t I and is a function of Therefore, the humidity ratio of an air-water vapor mixture can be detennined from the entering and leaving tempera­ tures and pressures '(If the adiabatic saturator. Consider the following example.

i:

t;.

The pressure entering and leaving an adiabatic saturator is 14.6% Ibflin.2, the entering temperature is 80 F, and the leaving temperature is 64 F. Compute the humidity ratio WI and the relative humidity 4>1'

SOLUTION Because the mixture leaving the device is saturated, P v2 = P.2, and Wz can be calculated using Eq. 3-14b.

l,1

3-4

Wet Bulb Temperature and the Psychrometric Chart

61

I 'l

~i

1 l

'I

I !

!l ~

I ~~

by the psychrometer are called the wet bulb and the dry bulb temperatures. The dry bulb temperature corresponds to t I in Fig. 3-1 and the wet bulb temperature is an approxima­ tion to in Fig. 3-1, whereas PI and P2 are equal to barometric or total mixture pressure. The combination heat-and-mass-transfer process from the wet bulb thermometer is not the same all the adiabatic saturation process; however, the error is relatively small when the wet bulb thermometer is used under suitable conditions. When the wet bulb temperature appears in psychrometric equations and charts, it is really the adiabatic saturation temperature that is being considered. Threlkeld (4) has analyzed the problem and correlated wet bulb temperature with the adiabatic saturation temperature. The results are shown in Figs. 3-3 and 3-4 for thermometers with diameters of 0.3 and 0.1 in. or 7.5 and 2.5 mm. The data are shown as percent deviation of (twb - t;> from the wet bulb depression (t - t wb) for shielded and unshielded wet bulbs. Three different temperature combinations are given and several general conclusions may be drawn. It appears that the unshielded wet bulb will generally more closely approximate the adiabatic saturation temperature. Temperatures taken with air velOCities greater than about 200 ftlmin (1 mls) generally have the least deviation. Lower air velocities may be used with smaller wet bulbs. Threlkeld drew the following general conclusion: For atmospheric temperature above freezing. where the wet bulb depression does not exceed about 20 F (11 C), and where no unusual radiation

t;

I

Air velocitv, mI.

~

o 6,0 .---.--=;~

f

4.0 I

I I -I­

14

16

1 1 I

I

'j"'nu;u Wt;f uu',":

J

!

~,

l

1 J ~

i1 .!, >. I,

2.0

I \\:1

1 1 1 1 I

1 1 1 I

I

I

I tJ I !S\ ~ a I I I 1-t:C ~It -'"W-t3ttl-1 1 .1 rtt!] I g

~:r:r:::o

t=120Ft50C).t"'6=I00Ft38C)

*1

o

-4.0 f-I-I--!----"I--;

~

I

I

i

1 :1

.~

-6,01

1 ......1..... 1

o

I, 1600

2000

2400

2800

3200

Air velocity. ft/min

Figure 3-3 Deviation (twb - t*) in percent of the,wet bulb depression (t - twb ) for a wet bulb diameter of 0.3 in. (7.5 mm) and a barometric pressure of 14.696 psia (101.325 lcPa). (James L. Threlkeld, Thermal Environmental Engineering, 2nd ed. 1910, p. 208. Reprinted by permission of Prentice-Hall, Inc., Engle-, wood Cliffs, NJ).

62

Chapter 3

Moist Air Properties and Conditioning Processes

6.0

o

6

4

2

!~

Air velocity, mI. 8 10

12

14

..

---I--- 1----- I--­

- - Unshielded wet bulb (t. - d - - - Shielded wet bulb

4.0

16

2.0

8

~

.r-., ~

I

~

~

0

... 1

~~

--2.0

\

-4.0

,l--­

,-

~(t

1-,-~ ~

1-­

.-

"

~

/'

o

I

400

I-

800

.­

It = 90 FI33 C), t..b= 70 F (21 C)

It - 30 F I 1 ct, t"~ = 20 F

--

V

I ~

f-­

I'-­

-6.0

120 F (SO C), tUth= 100 F (38 C)

/'

/ IJ­

-

,I

7 C)

1-- ._->--1""-

.­ I

1200 1600 2000 Air velocity, H/min

2400

2800

3200

Figure 3-4 Deviation (t",p - t*) in percent of the wet bulb depression (1 twtJ for a wet bulb diameter of 0.1 in. (2.5 mm) and a barometric pressure of 14.696 psia (101.325 kPa). (James L. Threlkeld. Thermal Environmental Engineering. 2nd erl. 1970, p. 209. Reprinted by pennission of Prentice-Hall, Inc., Engle­ wood Cliffs, NJ)

circumstances exist, (twb - t;> should be less than about 0.5 F (0.27 C) for an unshielded mercury-in-glass thermometer as long as the air velocity exceeds about lqo ftlmin (0.5 mls). If thermocouples are used, the velocity may be somewhat lower with similar accuracy. Figure 3-5 shows a psychrometer properly installed to meet the foregoing conditions.

Wick

Pressure equalizing tube and water reservoir

Figure 3-5 A psychrometer installed in a duct.

3-4

~

j

i

! ~

1

'1. . l

I

.

~ l

i t

1

I1

Wet Bulb Temperature and the Psychrometric Chart

63

Thus, for most engineering problems the wet bulb temperature obtained from a properly operated, unshielded psychrometer may be used directly in Eq. 3-2ld in place of the adiabatic saturation temperaJure. To facilitate engineering computations, a graphical representation of the properties of moist air has been developed and is known as a psychrometric chart. Richard Mollier was the first to use such a chart with enthalpy as a coordinate. Modem-day charts are somewhat different but· still retain the enthalpy coordinate feature. ASHRAE has developed five of the Mollier-type charts to cover the necessary range of variables. Figure 3-6 is an abridgment of ASHRAE Chart I that covers the normal range of variables at standard atmospheric pressure. The charts are based on the precise data of Tables A-2; within the readability of the charts, however, agreement with the perfect gas relations is very good. Details of the actual construction of the charts may be found in references 3 and 6. Charts for both English and SI units are provided in the packet in the back of the book. In Fig. 3-6 dry bulb temperature is plotted along the horizontal axis. The dry bulb temperature lines are straight but not exact;ly parallel and incline slightly to the left. Humidity ratio is plotted along the vertical axis on the right-hand side of the chart. The scale is uniform with horizontal lines. The saturation curve slopes upward from left to right. Dry bulb, wet bulb, and dew point temperatures all coincide on the saturation curve. Relative humidity lines with shapes similar to the saturation curve appear at regular intervals. The enthalpy scale is drawn obliquely on the left of the chart with parallel enthalpy lines inclined downward to the right. Although the wet bulb tempera­ ture lines appear to coincide with the enthalpy lines, they diverge gradually in the body of the chart and are not parallel to one another. The spacing of the wet bulb lines is not uniform. Specific volume lines appear inclined from the upper left to the lower right and are not parallel. A protractor with two scales appears at the upper left of ASHRAE Charts 1. One scale gives the sensible heat ratio and the other the ratio of enthalpy difference to humidity ratio difference. The enthalpy, specific volume, and humidity ratio scales are all based on a unit mass ojdry air and not a unit mass ojthe moist air.

!•

1

1

~

~

Read the properties of moist air at 75 F db, 60 F wb. and standard sea level pressure . from ASHRAE Psychrometric Chart lao

SOLUTION The intersection of the 75 F db and 60 F wb lines defines the given state. This point on the chart is the reference from which all the other properties are determined.

= 0.OO771bmvnbma on the vertical scale. Relative Humidity, ~ Interpolate between the 40 and 50 percent relative humidity lines and read q, = 41 percent. Enthalpy, i. Follow a line of constant enthalpy upward to the left and read i 26.4 Btunbma on the oblique scale. Specific Volume, v. Interpolate between the 13.5 and 14.0 specific volume lines and read v = 13.65 ft3nbma. Humidity Ratio, W. Move horizontally to the right and read W

...

0\

(J

:r

~ j

Ii

i....

w

~

9,

~ooO

~ -.;..!­

It. Blpy "'11; ,

14

/3

1:;.'

a-: .... '"Cl

(I'ly rBtio - /j.W

(3

"0 (1)

ij?

::l

E'

'"., ::l

~

$

(J 0

is.

g:

::l

S'

(lQ

$

"0

(3

(') (1)

'"'"~ ~

3-5

Classic Moist Air Processes

65

Dew Point Temperature, lifo Move horizontally to the left from the reference point and read td = 50 F on the saturation curve. Enthalpy, i (alternate method). The nomograph in the upper left-hand comer of Chart la gives the difference D between the enthalpy of unsaturated moist air and enthalpy of saturated air at the same wet bulb temperature. Then i = is + D. For this example is 26.5 Btu/lbma, D '" -0.1 Btullbma, and i = 26.5 - 0.1 26.4 Btullbma.

I

I I ~

Although psychrometric charts are useful in several aspects of HVAC design, the availability of computer programs to determine moist air properties have made so~e of these steps easier to carry out. These programs may be easily constrocted from the basic equations of this chapter. Computer programs give the additional convenience ofchoice of units and arbitrary mixture pressures.

3-5 CLASSIC MOIST AIR PROCESSES Two powerful analytical tools of the HVAC design engineer are the first law of thermodynamics or energy balance, and the conservation of mass or mass balance. These conservation laws are the basis for the analysis of moist air processes. In actu:iJ practice the properties may not be uniform across the flow area especially at the outlet, and a considerable length may be necessary for complete mixing. It is customary to analyze these processes by using the bulk average properties at the inlet and outlet ofthe device being studied. In this section we will consider the basic processes that are a part of the analysis of most systems.

Heating or Cooling of Moist Air

I

I ~ !!

When air is heated or cooled without the loss or gain of moisture, the process yields a straight horizontal line on the psychrometric chart because the humidity ratio is constant. Such processes can occur when moist air flows through a heat exchanger. In cooling, if part of the surface of the heat exchanger is below the dew point temperature of the air, condensation and the consequent dehumidification will occur. Figure 3-7 shows a schematic of a.device used to heat or cool air. Under steady-flow-steady-state conditions the energy balance becomes

i

! ~'. 8

mai2 +

~

it

!

2

= lOl + Wli"l

(3-23)

+ W2i Y2

(3·24)

and

ia

I

(3-22)

An energy balance yields a positive number for q for both cooling and heating, and the direction of the heat transfer is implied by the terms heating and cooling. The enthalpy of the moist air, per unit mass of dry air, at sections 1 and 2 is given by:

~

i

q = mail

i ez

Alternatively i I and 12 may be obtained directly from tlle psychrometric chart. The

66

Chapter 3

Moist Air Properties and Conditioning Processes Heating or cooling medium ~

~

~

~ 1'1',

~ ~

Figure 3-1 Schematic of a heating vice.

or cooling de­

1'1', = 1'1'2

It'

'--------L._L-------JI

-+-I db

I,

12

i

Figure 3·8 Sensible heating and cooling process.

convenience of the chart is evident. Figure 3-8 shows heating and cooling processes. Because the moist air has been assumed to be a perfect gas, Eq. 3-22 maybe arranged and written

iIs

m"citl-

(cooling and heating)

(3-25)

+

(3-26)

where Cp

=

Cpu

In the temperature range of interest, cpa = 0.24 Btul(lbma-F) or 1.0 kl/(lcga-C), cpu 0.45 Btul(lbmv-F) or 1.86 kl/(kga~C), and W is the order of 0.01. Then cp is about 0.245 Btul(lbma-F) or 1.02 kl/(kga-C).

Find the heat transfer rate required to warm 1500 cfrn (ft 3/min) of air at 60 F and 90 percent relative humidity to 120 F without the addition of moisture.

~

1

'l

3-5

.j

Classic Moist Air Processes

67

SOLUTION !

J "

Equation 3-22 or 3-25 may be used to find the required heat transfer rate. First it is necessary to find the mass flow rate of the dry air.

1 l

rna =

I

=

VIAl --;;-

(3-27)

lJ I

j

i

I

The specific volume is read from Chart la at tl

mil

l

J

=

1500(60) 13.31

I

Also from Chart la, i l we get

I

or if we had chosen to use Eq. 3-25

I,

67621bmalhr

25.3 Btullbma and ;2

q q=

6762(40.0

25.3)

40 Btullbma. Then by using Eq. 3-22, 99,400 Btulhr

6762(0.245)(120 - 60) = 99,400 Btulhr

Locating point 2 at the end of the heating process on the chart, we can see ,that the relative humidity decreases when the moist air is heated. The reverse process ofcooling results in an increase in relative humidity.

Cooling and Dehumidifying of Moist Air When moist air is cooled to a temperature below its dew point, some of the water vapor will condense and may leave the air stream. Figure 3-9 shows a schematic of a cooling and dehumidifying device, and Fig. 3-10 shows the process on the psychrometric chart. Although the actuafprocess path may vary considerably depending on the type of surface, surface temperature, and flow conditions, the net heat and mass transfer can be expressed in terms of the initial and final states. By referring to Fig. 3-9, we see that the energy balance gives

mail =

1

II

90 percent as 13.31 ft?1

60 F and '"

Ibrna:

if + m.,l2 + mwiw

(3·28)

41h ct' Refrigerant