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Full text of "Principles of Refrigeration"

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Wiley International Edition 



PRINCIPLES OF 
REFRIGERATION 

ROY J. DOSS AT, Associate Professor of Refrigeration 
and Air Conditioning, University of Houston, Houston, Texas 



W 



JOHN WILEY & SONS, INC. 
NEW YORK and LONDON 

TOPPAN COMPANY, LTD. 
TOKYO, JAPAN 



Authorized reprint of the edition published by John 
Wiley & Sons, Inc., New York and London. 

Copyright © 1961 by John Wiley ft Sons, Inc. 

All Rights Reserved. No part of this book may 
be reproduced in any form without the written 
permission of John Wiley & Sons, Inc. 

Wiley International Edition 

This book is not to be sold outside the country 
to which it is consigned by the publisher. 

Library of Congress Catalog Card Number: 61-15396 

Printed in Singapore by Toppan Printing Co. (S) Pte. Ltd. 



Preface 



This textbook has been written especially for use in programs where a full 
curriculum in refrigeration is offered. However, the material covered and the 
method of presentation are such that the text is also suitable for adult evening 
classes and for on-the-job training and self-instruction. Furthermore, the 
material is so arranged and sectionalized that this textbook is readily adaptable 
to any level of study and to any desired method or sequence of presentation. 

Despite a rigorous treatment of the thermodynamics of the cycle, application 
of the calculus is not required nor is an extensive background in physics and 
thermodynamics presupposed. The first four chapters deal with the funda- 
mental principles of physics and thermodynamics upon which the refrigeration 
cycle is based. For those who are already familiar with these fundamentals, 
the chapters will serve as review or reference material. 

Chapter 21 treats electric motors and control circuits as they apply to refrigera- 
tion and air conditioning systems. This material is presented from the viewpoint 
of practical application, the more mathematical approach being left to companion 
electrical courses. 

Throughout this textbook emphasis is placed on the cyclic nature of the 
refrigeration system, and each part of the system is carefully examined in relation 
to the whole. Too, care is taken continually to correlate theory and practice 
through the use of manufacturer's catalog data and many sample problems. To 
this end, certain pertinent catalog data are included. 

Roy J. Dossat 

July, 1961 



Acknowledgments 



Most of the material in this textbook 
is based on information gathered from 
publications of the American Society of 
Heating, Refrigerating, and Air Condition- 
ing Engineers and of the following equip- 
ment manufacturers : 

Acme Industries, Inc. 
Alco Valve Company 
Anaconda Metal Hose Division, The 

American Brass Company 
Bell & Gossett Company 
Carrier Corporation 
Controls Company of America 
Dean Products, Inc. 
Detroit Controls Division, American 

Radiator & Standard Sanitary 

Corporation 
Detroit Ice Machine Company 
Dole Refrigerating Company 
Dunham-Bush, Inc. 
Edwards Engineering Corporation 
E. I. du Pont de Nemours & Company 
Freezing Equipment Sales, Inc. 
Frick Company 
General Controls Company 
General Electric Company 
Halstead & Mitchell 



Ingersoll-Rand Company 

Kennard Division, American Air Filter 

Company, Inc. 
Kramer Trenton Company 
McQuay, Inc. 
The Marley Company 
Marsh Instrument Company 
Mueller Brass Company 
Penn Controls, Inc. 
Recold Corporation 
Sporlan Valve Company 
Tecumseh Products Company 
Tranter Manufacturing, Inc. 
Tubular Exchanger Manufacturers 

Association, Inc. 
Tyler Refrigeration Corporation 
The Vilter Manufacturing Company 
worthington corporation 
York Corporation, Subsidiary of Borg- 

Warner Corporation 

Appreciation is expressed to all these 
organizations for their contributions in the 
form of photographs and other art work, 
and for granting permission to reproduce 
proprietary data, without which this text- 
book would not have been possible. 



vi 



Contents 



1. Pressure, Work, Power, Energy I 

2. Matter, Internal Energy, Heat, Temperature 10 

3. Thermodynamic Processes 24 

4. Saturated and Superheated Vapors 43 

5. Psychrometric Properties of Air 57 

6. Refrigeration and the Vapor Compression System 71 

7. Cycle Diagrams and the Simple Saturated Cycle 89 

8. Actual Refrigerating Cycles 107 

9. Survey of Refrigeration Applications 121 

10. Cooling Load Calculations 144 

11. Evaporators 164 

12. Performance of Reciprocating Compressors 203 

13. System Equilibrium and Cycling Controls 225 

14. Condensers and Cooling Towers 244 

15. Fluid Flow, Centrifugal Liquid Pumps, Water and Brine Piping 274 

16. Refrigerants 284 

17. Refrigerant Flow Controls 298 



vii 



viii PRINCIPLES OF REFRIGERATION 

18. Compressor Construction and Lubrication 334 

19. Refrigerant Piping and Accessories 365 

20. Defrost Methods, Low-Temperature Systems, and Multiple Temperature 
Installations 388 

21. Electric Motors and Control Circuits 407 
Tables and Charts 430 

Index 535 



I 



Pressure, Work, 
Power, Energy 



l-l. Force. A force is defined as a push or a 
pull. It is anything that has a tendency to set 
a body in motion, to bring a moving body to 
rest, or to change the direction of motion. A 
force may also change the size or shape of a 
body. That is, the body may be twisted, bent, 
stretched, compressed, or otherwise distorted 
by the action of a force. 

The most familiar force is weight. The 
weight of a body is a measure of the force 
exerted on the body by the gravitational pull 
of the earth (Fig. 1-1). 

There are many forces other than the force of 
gravity, but all forces are measured in weight 
units. Although the most commonly used unit 
of force measure is the pound, any unit of 
weight measure may be used, and the particular 
unit- used at any time will usually depend on 
the magnitude of the force to be measured. 
1-2. Pressure. Pressure is the force exerted 
per unit of area. It may be described as a 
measure of the intensity of a force at any given 
point on the contact surface. Whenever a force 
is evenly distributed over a given area, the 
pressure at any point on the contact surface is 
the same and can be calculated by dividing the 
total force exerted by the total area over which 
the force is applied. This relationship is 
expressed by the following equation 



where P «■ the pressure expressed in units of F 
per unit of A 
F — the total force in any units of force 
A = the total area in any units of area 
1-3. Measurement of Pressure. As indi- 
cated by equation 1-1, pressure is measured 
in units of force per unit of area. Pressures are 
most frequently given in pounds per square 
inch, abbreviated psi. However, pressure, like 
force, as a matter of convenience and depending 
on the magnitude of the pressure, may be stated 
in terms of other units of force and area, such 
as pounds per square foot, tons per square foot, 
grams per square centimeter, etc. 

Example l-l. A rectangular tank, measur- 
ing 2 ft by 3 ft at the base, is filled with water. 
If the total weight of the water is 432 lb, deter- 
mine the pressure exerted by the water on the 
bottom of the tank in 

(a) pounds per square foot 

(b) pounds per square inch 

Solution 

(a) Area of tank base 



Total weight of water 



'-h 



(1-1) 



2x3 

= 6sqft 

= 432 lb 

432 
Applying Equation 1-1, P —^ 

= 72psf 
(b) Area of the tank base 24 x 36 

= 864 sq in. 
Total weight of water = 432 lb 

432 
Applymg Equation 1-1, P 

864 
= 0.5 psi 

The problem described in Example 1-1 is 
illustrated in Fig. 1-2. Notice that the pressure 
on the bottom of the tank in pounds per square 
foot is equivalent to the downward force 
exerted by the weight of a column of water 
having a cross section of one square foot, where- 
as the pressure in pounds per square inch is 
equivalent to the downward force exerted by a 
column of water having a cross section of 1 sq in. 
Further, since there are 144 sq in. in 1 sq ft, the 
force exerted per square foot is 1 44 times as great 
as the force exerted per square inch. 
1-4. Atmospheric Pressure. The earth is 
surrounded by an envelope of atmosphere or air 
which extends upward from the surface of the 
earth to a distance of some 50 mi or more. Air 
has weight and, because of its weight, exerts a 



PRINCIPLES OF REFRIGERATION 



Pointer - 



-Spring scale 




Weight 



Fig. 1-1, Because of gravity, the suspended weight 
exerts a\downward force of 7 lb. 

pressure on the surface of the earth. The pressure 
exerted by the atmosphere is known as atmos- 
pheric pressure. 

The weight of a column of air having a cross 
section of 1 sq in. and extending from the surface 
of the earth at sea level to the upper limits of the 
atmosphere is 14.696 lb. Therefore, the pressure 
on the surface of the earth at sea level resulting 
from the weight of the atmosphere is 14.696 psi 
(14.7). This is understood to be the normal or 
standard atmospheric pressure at sea level and 
is sometimes referred to as a pressure of one 
atmosphere. Actually, the pressure of the 
atmosphere does not remain constant, but will 
usually vary somewhat from hour to hour 
depending upon the temperature, water vapor 
content, and several other factors. 

Because of the difference in the height of the 
column, the weight of a column of air of given 



cross section will be less when taken at an 
altitude of one mile above sea level than when 
taken at sea level. Therefore, it follows that 
atmospheric pressure decreases as the altitude 
increases. 

1-5. Barometers. Barometers are instruments 
used to measure the pressure of the atmosphere 
and are of several types. A simple barometer 
which measures atmospheric pressure in terms 
of the height of a column of mercury can be 
constructed by filling with mercury a hollow 
glass tube 36 in. or more long and closed at one 
end. The mercury is held in the tube by placing 
the index finger over the open end of the tube 
while the tube is inverted in an open dish of 
mercury. When the finger is removed from the 
tube, the level of the mercury in the tube will 
fall, leaving an almost perfect vacuum at the 
closed end. The pressure exerted downward by 
the atmosphere on the open dish of mercury will 
cause the mercury to stand up in the evacuated 
tube to a height depending upon the amount of 
pressure exerted. The height of the mercury 
column in the tube is a measure of the pressure 
exerted by the atmosphere and is read in inches 
of mercury column (abbreviated in. Hg). The 
normal pressure of the atmosphere at sea level 
(14.696 psi) pressing down on the dish of mercury 
will cause the mercury in the tube to rise to a 




Fig. 1-2. Of the total weight of the water in the 
container, that part which is exerted on a I sq ft area 
is the pressure in pounds per square foot. Likewise, 
that part which is exerted on a I sq in. area is the 
pressure in pounds per square inch. 



PRESSURE, WORK, POWER, ENERGY 3 



height of 29.921 in. (Fig. 1-3). A column of 
mercury 29.921 in. high is, then, a measure of a 
pressure equivalent to 14.696 psi. By dividing 
29.921 in. Hg by 14.696 psi, it is determined that 
a pressure of 1 psi is equivalent to a pressure of 
2.036 in. Hg. Therefore, 1 in. Hg equals 
1/2.036, or 0.491 psi, and the following equa- 
tions are established: 

psi 



in. Hg = ■ 



(1-2) 



0.491 

and psi = in. Hg x 0.491 (1-3) 

Example 1-2. What is the pressure of the 
atmosphere in psi if a barometer reads 30.2 
in. Hg? 

Solution. Applying Equation 1-3, 
P - 30.2 x 0.491 
= 14.83 psi 
Example 1-3. In Fig. 1-3, how high will 
the mercury stand in the tube when the atmos- 
pheric pressure is 14.S psi? 

Solution. Applying Equation 1-2, 

0.491 
= 29.53 in. Hg 

1-6. Pressure Gages. Pressure gages are 
instruments used to measure the fluid pressure 
(either gaseous or liquid) inside a closed vessel. 
Pressure gages commonly used in the refriger- 
ation industry are of two types: (1) manometer 
and (2) bourdon tube. 

1-7. Manometers. The manometer type gage 
utilizes a column of liquid to measure the 
pressure, the height of the column indicating 
the magnitude of the pressure. The liquid used 
in manometers is usually either water or mercury. 
When mercury is used, the instrument is known 
as a mercury manometer or mercury gage and, 
when water is used, the instrument is a water 
manometer or water gage. The simple barom- 
eter described previously is a manometer type 
instrument. 

A simple mercury manometer, illustrated in 
Figs. 1-4a, 1-46 and l-4c, consists of a U-shaped 
glass tube open at both ends and partially filled 
with mercury. When both legs of the U-tube 
are open to the atmosphere, atmospheric pres- 
sure is exerted on the mercury in both sides of 
the tube and the height of the two mercury 
columns is the same. The height of the two 
mercury columns at this position is marked as 
the zero point of the scale and the scale is cali- 



brated in inches to read the deviation of the 
mercury columns from the zero condition in 
either direction (Fig. l-4a). 

When in use, one side of the U-tube is 
connected to the vessel whose pressure is to be 
measured. The pressure in the vessel, acting on 
one leg of the tube, is opposed by the atmos- 
pheric pressure exerted on the open leg of the 
tube. If the pressure in the vessel is greater 
than that of the atmosphere, the level of the 
mercury on the vessel side of the U-tube is 
depressed while the level of the mercury on the 
open side of the tube is raised an equal amount 
(Fig. l-4b). If the pressure in the vessel is less 
than that of the atmosphere, the level of the 
mercury in the open leg of the tube is depressed 
while the level of the mercury in the leg con- 
nected to the vessel is raised by an equal amount 
(Fig. l-4c). In either case, the difference in the 
heights of the two mercury columns is a 
measure of the difference in pressure between 
the total pressure of the fluid in the vessel and 
the pressure of the atmosphere. 

In Fig. l-4b, the level of the mercury is 2 in. 




Scale (inches) 



Dish of mercury 



Fig. 1-3. The pressure exerted by the weight of the 
atmosphere on the open dish of mercury causes the 
mercury to stand up into the tube. The magnitude 
of the pressure determines the height of the mercury 
column. 



PRINCIPLES OF REFRIGERATION 



Atmospheric 
/pressure \ 




Fig. l-4o. Simple U-tube manometer. Since both 
legs of the manometer are open to the atmosphere 
and are at the same pressure, the level of the mercury 
is the same in both sides. 



Atmospheric 
pressure 
30 in. Hg 




Fig. I -4b. Simple manometer indicates that the 
vessel pressure exceeds the atmospheric pressure 



vessel pressure 
by 4 in. Hg. 

below the zero point in the side of the U-tube 
connected to the vessel and 2 in. above the zero 
point in the open side of the tube. This indi- 
cates that the pressure in the vessel exceeds the 
pressure of the atmosphere by 4 in. Hg (1.96 
psi). In Fig. l-4c, the level of the mercury is 
depressed 2 in. in the side of the tube open to 
the atmosphere and raised 2 in. in the side con- 
nected to the vessel, indicating that the pressure 
in the vessel is 4 in. Hg (1.96 psi) below (less 
than) atmospheric. Pressures below atmospheric 
are usually called "vacuum" pressures and may 
be read as "inches of mercury, vacuum." 



Manometers using water as the measuring 
fluid are particularly useful for measuring very 
small pressures. Because of the difference in the 
density of mercury and water, pressures so 
slight that they will not visibly affect the height 
of a mercury column will produce easily 
detectable variations in the height of a water 
column. Atmospheric pressure, which will 
support a column of mercury only 29.921 in. 
high, will lift a column of water to a distance of 
approximately 34 ft. A pressure of 1 psi will 
raise a column of water 2.31 ft or 27.7 in. and a 



Atmospheric 
pressure 
30 in. Hg 




Vessel 

- pressure 

26 in. Hg 



Fig. l-4c Manometer indicates that the vessel 
pressure is 4 in. Hg less than the atmospheric 
pressure of 30 in. Hg. 




Fig. 1-5. Bourdon tube gage mechanism. (Courtesy 
Marsh Instrument Company.) 



pressure of only 0.036 psi is sufficient to support 
a column of water 1 in. high. Hence, 1 in. 
of water column is equivalent to 0.036 psi. 

Table 1-1 gives the relationship between the 
various units of pressure measurement. 
1-8. Bourdon Tube Gages. Because of the 
excessive length of tube required, gages of the 
manometer type are not practical for measuring 
pressures above IS psi and are more or less 



PRESSURE, WQRK, POWER, ENERGY 5 

inches of mercury (Fig. 1-66). In many cases, 
single gages, known as "compound" gages, are 
designed to measure pressures both above and 
below atmospheric (Fig. l-6c). Such gages are 
calibrated to read in psi above atmospheric and 
in inches of mercury below atmospheric. 
1-9. Absolute and Gage Pressures. Absolute 
pressure is understood to be the "total" or 
"true" pressure of a fluid, whereas gage pressure 






<°) 



(b) 



(c) 



Fig. 1-6. Typical bourdon tube gages, (a) Pressure gage, (b) Vacuum gage, (c) Compound gage. (Courtesy 
Marsh instrument Company.) 



limited to the measurement of relatively small 
pressures in air ducts, etc. 

Gages of the bourdon tube type are widely 
used to measure the higher pressures en- 
countered in refrigeration work. The actuating 
mechanism of the bourdon tube gage is illus- 
trated in Fig. 1-5. The bourdon tube, itself, is a 
curved, elliptical-shaped, metallic tube which 
tends to straighten as the fluid pressure in the 
tube increases and to curl tighter as the pressure 
decreases. Any change in the curvature of the 
tube is transmitted through a system of gears 
to the pointer. The direction and magnitude 
of the pointer movement depend on the direc- 
tion and magnitude of the change in the curva- 
ture of the tube. 

Bourdon tube gages are very rugged and will 
measure pressures either above or below 
atmospheric pressure. Those designed to 
measure pressures above atmospheric are known 
as "pressure" gages (Fig. l-6a) and are gener- 
ally calibrated in psi, whereas those designed to 
read pressures below atmospheric are called 
"vacuum" gages and are usually calibrated in 



is the pressure as indicated by a gage. It is 
important to understand that gages are cali- 
brated to read zero at atmospheric pressure and 
that neither the manometer nor the bourdon 
tube gage measures the "total" or "true" 
pressure of the fluid in a vessel; both measure 
only the difference in pressure between the total 
pressure of the fluid in the vessel and the atmos- 
pheric pressure. When the fluid pressure is 
greater than the atmospheric pressure, the 
absolute pressure of the fluid in the vessel is 
determined by adding the atmospheric pressure 
to the gage pressure, and, when the fluid 
pressure is less than atmospheric, the absolute 
pressure of the fluid is found by subtracting the 
gage pressure from the atmospheric pressure. 
The relationship between absolute pressure and 
gage pressure is shown graphically is Fig. 1-7. 

Example 1-4. A pressure gage on a 
refrigerant condenser reads 120 psi. What is 
the absolute pressure of the refrigerant in the 
condenser? 

Solution. Since the barometer reading is not 
given, it is assumed that the atmospheric 



PRINCIPLES OF REFRIGERATION 



Gage 
pressure 



45- 
40 
35- 
30 
25- 
20- 
15- 
10- 
5- 



Absolute 
pressure 



Pressures above 
atmospheric in psi 



Atmospheric pressure 



-Pressures below- 
- atmospheric in— 
in. Hg 



59.7 
54.7 
49.7 

-44.7 
39.7 
34.7 
29.7 
24.7 

•19.7 



29.92 in. Hg 
(14.7 psi) 



5- 
10 
15- 
20 

(14.7 psi) 25 
29.92 in. Hg~~ 



Fig. 1-7. Relationship between absolute and gage 
pressures. 

pressure is normal at sea level, 14.696 psi, and, 
since the pressure of the refrigerant is above 
atmospheric, the absolute pressure of the 
refrigerant is equal to the gage pressure plus 
the atmospheric pressure. 

Gage pressure in psi =120 

Atmospheric pressure in psi = 14.696 

Absolute pressure of 
refrigerant = 134.696 psi 

Example 1-5. A compound gage on the 
suction side of a vapor compressor reads 
5 in. Hg, whereas a barometer nearby reads 
29.6 in. Hg. Determine the absolute pressure 
of the vapor entering the compressor. 

Solution. Since the pressure of the vapor 
entering the compressor is less than atmospheric, 
the absolute pressure of the vapor is computed 
by subtracting the gage pressure from atmos- 
pheric pressure. 

Atmospheric pressure in 
in. Hg = 29.6 

Gage pressure in in. Hg = 5.0 

Absolute pressure in 
in. Hg = 24.6 in. Hg 

Absolute pressure 24.6 x 0.491 

= 12.08 psi 



Example 1-6. During compression the 
pressure of a vapor is increased from 10 in. Hg 
gage to 125 psi gage. Calculate the total increase 
in pressure in psi. 

Solution. Since the pressure increases from 
10 in. Hg below atmospheric to 125 psi above 
atmospheric, the total increase in pressure is 
the sum of the two pressures. 

Initial pressure = 10 in. Hg 
Initial pressure in psi below 1 x 0.491 

atmospheric = 4.91 psi 
Final pressure in psi above 

atmospheric =125 psi 

Total increase in pressure = 129.91 psi 
Absolute pressure in psi is abbreviated psia, 
whereas gage pressure in psi is abbreviated psig. 
1-10. Work. Work is done when a force 
acting on a body moves the body through a 
distance. The amount of work done is the 
product of the force and the distance through 
which the force acts. This relationship is shown 
by the following equation: 

W = F x 1 (1-4) 

where F = the force applied in any units of 
force 
1 = the distance through which the force 

acts in any linear unit 
W = the work done expressed in units of 
force and linear measure 

The work done is always expressed in the 
same unit terms used to express the magnitude 
of the force and the distance. For instance, if 
the force is expressed in pounds and the dis- 
tance in feet, the work done is expressed in 
foot-pounds. The foot-pound is the most 
frequently used unit of work measure. 

Example U7. A ventilating fan weighing 
315 lb is hoisted to the roof of a building 200 ft 
above the level of the ground. How much 
work is done? 



Solution. By applying 
Equation 1-4, the weight of 
the fan 

Distance through which 
the fan is hoisted 

Work done 



= 315 lb 

= 200 ft 

315 x 200 
= 63,000 ft-lb 



I -I I. Power. Power is the rate of doing 
work. That is, it is the work done divided by the 
time required to do the work. The unit of power 



PRESSURE, WORK, POWER, ENERGY 7 



is the horsepower. One horsepower is defined as 
the power required to do work at the rate of 
33,000 ft-lb per minute or (33,000/60) 550 ft-lb 
per second. The power required in horsepower 
may be found by either of the following equa- 
tions: 

^-TSMTT, (1 " 5) 

where Hp = the horsepower 

W = the work done in foot-pounds 



t = the time in minutes 



or 



Hp = 



W 



(1-6) 



550 x t 

where t = the time in seconds 

Example 1-8. In Example 1-7, if the time 
required to hoist the fan to the roof of the 
building is 5 minutes, how much horsepower 
is required? 

Solution. Total work done =63,000 ft-lb 

Time required to do the 

work = 5 min 

„ . J 63,000 

Horsepower required 33,000 x 5 

= 0.382 hp 
1-12. Energy. In order to do work or to 
cause motion of any kind, energy is required. 
A body is said to possess energy when it has 
the capacity for doing work. Hence, energy is 
described as the ability to do work. The 
amount of energy required to do a given amount 
of work is always equal to the amount of work 
done and the amount of energy a body possesses 
is equal to the amount of work a body can do 
in passing from one condition or position to 
another. 

Energy may be possessed by a body in either 
or both of two basic kinds: (1) kinetic and (2) 
potential. 

1-13. Kinetic Energy. Kinetic energy is the 
energy a body possesses as a result of its motion 
or velocity. For instance, a hammer swinging 
through an arc, a bullet speeding toward a 
target, and the moving parts of machinery all 
have kinetic energy by virtue of their motion. 
The amount of kinetic energy a body possesses 
is a function of its mass and its velocity and may 
be determined by the following equation: 

M x V 2 



where JT — the kinetic energy in foot-pounds 
M = the weight of the body in pounds 
V = the velocity in feet per second (fps) 
g = gravitational constant (32. 1 74 ft/sec 2 ) 

Example 1-9. An automobile weighing 
3500 lb is moving at the rate of 30 mph. What 
is its kinetic energy? 



Solution. Velocity 
in fps V 

Applying Equa- 
tion 1-7, the kinetic 
energy AT 



5280 ft/mi x 30 mi 
3600sec/hr 

= 44 fps 

3500 lb x (44 fps)* 
2 x 32.174 ft/sec 2 

= 105,302 ft-lb 



1-14. Potential Energy. Potential energy is 
the energy a body possesses because of its 
position or configuration. The amount of work 
a body can do in passing from a given position 
or condition to some reference position or 
condition is a measure of the body's potential 
energy. For example, the driving head of a pile- 
driver has potential energy of position when 
raised to some distance above the top of a 
piling. If released, the driving head can do the 
work of driving the piling. A compressed steel 
spring or a stretched rubber band possesses 
potential energy of configuration. Both the 
steel spring and the rubber band have the 
ability to do work because of their tendency to 
return to their normal condition. 

The potential energy of a body may be evalu- 
ated by the following equation: 



P =M xZ 



(1-8) 



where P = the potential energy in foot-pounds 
M = the weight of the body in pounds 
Z = the vertical distance above some 
datum or reference 

Example 1-10. Ten thousand gallons of 
water are stored in a tank located 250 ft above 
the ground. Determine the potential energy of 
the water in relation to the ground. 



Solution. The 
weight of the 
water in pounds 
M 

Applying 
Equation 1-8, 
the potential 
energy P 



10,000 gal x 8.33 lb/gal 
83,300 lb 



83,300 lb x 250 ft 
20,825,000 ft-lb 



8 PRINCIPLES OF REFRIGERATION 



1-15. Energy as Stored Work. Before a 
body can possess energy, work must be done on 
the body. The work which is done on a body 
to give the body its motion, position, or con- 
figuration is stored in the body as energy. 
Hence, energy is stored work. For instance, 
work must be done to stretch therubber band, to 
compress the steel spring, or to raise the driving 
head of a pile-driver to a position above the 
piling. In any case, the potential energy stored 
is equal to the work done. 

The amount of energy a body possesses can be 
ascertained by determining the amount of work 
done on the body to give the body its motion, 
position, or configuration. For example, assume 
that the driving head of a pile-driver weighing 
200 lb is raised to a position 6 ft above the top 
of a piling. The work done in raising the driving 
head is 1200 ft-lb (200 lb x 6 ft). Therefore, 
1200 ft-lb of energy are stored in the driving- 
head in its raised position and, when released, 
neglecting friction, the driving-head will do 
1200 ft-lb of work on the piling. 
1-16. Total External Energy. The total 
external energy of a body is the sum of its 
kinetic and potential energies. 

Example l-l I. Determine the total external 
energy of an airplane weighing 10,000 lb and 
flying 6000 ft above the ground at a speed of 
300 mph. 

Solution. Apply- 
ing Equation 1-7, 

the kinetic energy 10,000 lb x (440 fps), 

K 2 x 32.174 ft/sec* 

= 30,086,436 ft-lb 

Applying 
Equation 1-8, the 10,000 lb x 6000 ft 

potential energy P - 60,000,000 ft-lb 

Adding, the 
total external 
energy = 90,086,436 ft-lb 

1-17. Law of Conservation of Energy. The 

First Law of Thermodynamics states in effect 
that the amount of energy is constant. None 
can be either created or destroyed. Energy is 
expended only in the sense that it is converted 
from one form to another. 
1-18. Forms of Energy* All energy can be 
classified as being of either of the two basic 
kinds, kinetic or potential. However, energy 
may appear in any one of a number of different 
forms, such as mechanical energy, electrical 



energy, chemical energy, heat energy, etc., and 
is readily converted from one form to another. 
Electrical energy, for instance, is converted into 
heat energy in an electric toaster, heater, or 
range. Electrical energy is converted into 
mechanical energy in electric motors, solenoids, 
and other electrically operated mechanical 
devices. Mechanical energy, chemical energy, 
and heat energy are converted into electrical 
energy in the generator, battery, and thermo- 
couple, respectively. Chemical energy is con- 
verted into heat energy in chemical reactions 
such as combustion and oxidation. These are 
only a few of the countless ways in which the 
transformation of energy can and does occur. 
There are many fundamental relationships 
which exist between the various forms of energy 
and their transformation, some of which are of 
particular importance in the study of refrigera- 
tion and are discussed in detail later. 

PROBLEMS 

1. The cooling tower on the roof of a building 
weighs 1360 lb when filled with water. If the 
basin of the tower measures 4 ft by 5 ft, what 
is the pressure exerted on the roof 

(a) in pounds per square foot? Ans. 68 psf. 
(£) in pounds per square inch? 

Ans. 0.472 psi. 

2. If the atmospheric pressure is normal at sea 
level and a gage on an R-12 condenser reads 
130 psi, what is the absolute pressure of the 
Freon in the condenser in pounds per square 
inch? Ans. 144.7 psia. 

3. What is the total force exerted on the top- of 
a piston if the area of the cylinder bore is 5 
sq in. and the pressure of the gas in the cylinder 
islSQpsi? Ans. 7501b. 

4. A barometer reads 10 in. Hg. What is the 
atmospheric pressure in psi? Ans. 4.91 psi. 

5. A barometer on the wall reads 29.6 in. Hg 
while a gage on the tank of an air compressor 
indicates 105 psi. What is the absolute pressure 
of the air in the tank in pounds per square foot? 

Ans. 119.53 psia. 

6. A gage on the suction inlet of a compressor 
reads 10 in. Hg. Determine the absolute pres- 
sure of the suction vapor in psi. Ans. 9.79 psia. 

7. A gage on the suction side of a refrigeration 
compressor reads 5 in. Hg. If a gage on the 
discharge side of the compressor reads 122 psi, 
what is the increase in pressure during the 
compression? Ans. 124.46 psi. 



PRESSURE, WORK, POWER, ENERGY 9 



8. An electric motor weighing 236 lb is hoisted 
to the roof of a building in 2 min. If the roof 
is 125 ft above the ground, 

(a) How much work is done? 

Ans. 29,500 ft-lb 

(b) Neglecting friction and other losses, what 
is tiie horsepower required? 

Ans. 0.447 hp 

9. Compute the kinetic energy of an automobile 
weighing 3000 lb and moving at a speed of 
75 mph. Ans. 567,188 ft-lb 

10. What is the total external energy of the 
automobile in Problem 9 if the automobile is 
traveling along a highway 6000 ft above 
sea level? Ans. 18,000,000 ft-lb 

11. What is the total potential energy of 8000 
gal of water confined in a tank and located 



a mean distance of 135 ft above the 
ground? Ans. 8,996,400 ft-lb 

12. Water in a river 800 ft above sea level is 
flowing at the rate of 5 mph. Calculate the sum 
of the kinetic and potential energies per pound 
of water in reference to sea level. 

Ans. 800.84 ft-lb 

13. A water pump delivers 60 gal per minute 
of water to a water tank located 100 ft above 
the level of the pump. If water weighs 8.33 lb 
per gallon and if the friction of the pipe and 
other losses are neglected, 

(a) How much work is done? 

Ans. 49,980 ft-lb 
(6) Compute the horsepower required. 

Ans. 1.5 hp. 



2 

Matter, Internal 
Energy, Heat, 
Temperature 



2-1. Heat. Heat is a form of energy. This 
is evident from the fact that heat can be con- 
verted into other forms of energy and that other 
forms of energy can be converted into heat. 
However, there is some confusion as to exactly 
what energy shall be termed heat energy. 
Popular usage has made the concept of heat as 
internal or molecular energy almost universally 
accepted. Because of this, referring to heat as 
internal energy is almost unavoidable at times. 
On the other hand, from a strictly thermo- 
dynamic point of view, heat is denned as energy 
in transition from one body to another as a 
result of a difference in temperature between the 
two bodies. Under this concept, all other 
energy transfers occur as work. Both these 
concepts of heat will evolve in this and the 
following chapters. The term heat will be used 
hereafter in this book in either sense. 
2-2. Matter and Molecules. Everything in 
the universe that has weight or occupies space, 
all matter, is composed of molecules. Mole- 
cules, in turn, are made up of smaller particles 
called atoms and atoms are composed of still 
smaller particles known as electrons, protons, 
neutrons, etc. The study of atoms and sub- 
atomic particles is beyond the scope of this 
book and the discussion will be limited for the 
most part to the study of molecules and their 
behavior. 



The molecule is the smallest, stable particle 
of matter into which a particular substance can 
be subdivided and still retain the identity of 
the original substance. For example, a grain 
of table salt (NaCl) may be broken down into 
individual molecules and each molecule will 
be a molecule of salt, the original substance. 
However, all molecules are made up of atoms, 
so that it is possible to further subdivide a 
molecule of salt into its component atoms. But, 
a molecule of salt is made up of one atom of 
sodium and one atom of chlorine. Hence, if a 
molecule of salt is divided into its atoms, the 
atoms will not be atoms of salt, the original 
substance, but atoms of two entirely different 
substances, one of sodium and one of chlorine. 

There are some substances whose molecules 
are made up of only one kind of atoms. The 
molecule of oxygen (0 2 ), for instance, is 
composed of two atoms of oxygen. If a mole- 
cule of oxygen is divided into its two component 
atoms, each atom will be an atom of oxygen, 
the original substance, but the atoms of oxygen 
will not be stable in this condition. They will 
not remain as free and separate atoms of oxygen, 
but, if permitted, will either join with atoms or 
molecules of another substance to form a new 
compound or rejoin each other to form again a 
molecule of oxygen. 

It is assumed that the molecules that make up 
a substance are held together by forces of 
mutual attraction known as cohesion. These 
forces of attraction that the molecules have for 
each other may be likened to the attraction that 
exists between unlike electrical charges or 
between unlike magnetic poles. However, 
despite the mutual attraction that exists between 
the molecules and the resulting influence that 
each molecule has upon the others, the mole- 
cules are not tightly packed together. There is a 
certain amount of space between them and they 
are relatively free to move about. The mole- 
cules are further assumed to be in a state of 
rapid and constant vibration or motion, the 
rate and extent of the vibration or movement 
being determined by the amount of energy 
they possess. 

2-3. Internal Energy. It has been pre- 
viously stated that energy is required to do work 
or to cause motion of any kind. Molecules, 
like everything else, can move about only if they 
possess energy. Hence, a body has internal 



10 



MATTER, INTERNAL ENERGY, HEAT, TEMPERATURE 1 1 



energy as well as external energy. Whereas 
a body has external mechanical energy because 
of its velocity, position, or configuration in 
relation to some reference condition, it also has 
internal energy as a result of the velocity, 
position, and configuration of the molecules of 
the materials which make up the body. 

The molecules of any material may possess 
energy in both kinds, kinetic and potential. 
The total internal energy of a material is the 
sum of its internal kinetic and potential energies. 
This relationship is shown by the equation 

U = K + P (2-1) 

where U = the total internal energy 
K = the internal kinetic energy 
P = the internal potential energy 

2-4. Internal Kinetic Energy. Internal kinetic 
energy is the energy of molecular motion or 
velocity. When heat energy flowing into a 
material increases the internal kinetic energy, 
the velocity or motion of the molecules is 
increased. The increase in molecular velocity is 
always accompanied by an increase in the 
temperature of the material. Hence, a material's 
temperature is, in a sense, a measure of the 
average velocity of the molecules which make 
up the material. The more kinetic energy the 
molecules have, the greater is their movement 
and the faster they move. The more rapid the 
motion of the molecules, the hotter is the 
material and the more internal kinetic energy 
the material has. It follows, then, that if the 
internal kinetic energy of the material is di- 
minished by the removal of heat, the motion of 
the molecules will be slowed down or retarded 
and the temperature of the material will be 
decreased. 

According to the kinetic theory if the removal 
of heat continues until the internal kinetic 
energy of the material is reduced to zero, the 
temperature of the material will drop to Abso- 
lute Zero (approximately —460° 10 and the 
motion of the molecules will cease entirely.* 

* It is now known that the energy is not zero at 
Absolute Zero. It is the disorganization (entropy) 
which diminishes to zero. Heat is sometimes 
defined as "disorganized energy." Both the energy 
and the disorganization decrease as the temperature 
decreases. However, the disorganization decreases 
faster than the energy and therefore diminishes 
to zero before the energy reaches zero. 



2-5. States of Matter. Matter can exist in 
three different phases or states of aggregation: 
solid, liquid, or a vapor or gas. For example, 
water is a liquid, but this same substance can 
exist as ice, which is a solid, or as steam, which 
is a vapor or gas. 

2-6. The Effect of Heat on the State of 
Aggregation. Many materials, under the 
proper conditions of pressure and temperature, 
can exist in any and all of the three physical 
states of matter. It will be shown presently 
that the amount of energy the molecules of the 
material have determines not only the tem- 
perature of the material but also which of the 
three physical states the material will assume at 
any particular time. In other words, the addi- 
tion or removal of heat can bring about a change 
in the physical state of the material as well as a 
change in its temperature. 

That heat can bring about a change in the 
physical state of a material is evident from the 
fact that many materials, such as metals will 
become molten when sufficient heat is applied. 
Furthermore, the phenomenon of melting ice 
and boiling water is familiar to everyone. Each 
of these changes in the physical state is brought 
about by the addition of heat. 
2-7. Internal Potential Energy. Internal 
potential energy is the energy of molecular 
separation or configuration. It is the energy the 
molecules have as a result of their position in 
relation to one another. The greater the degree 
of molecular separation, the greater is the inter- 
nal potential energy. 

When a material expands or changes its 
physical state with the addition of energy, a 
rearrangement of the molecules takes place 
which increases the distance between them. 
Inasmuch as the molecules are attracted to one 
another by forces which tend to pull them to- 
gether, internal work must be done in order to 
separate further the molecules against their 
attractive forces. An amount of energy equal to 
the amount of internal work done must flow 
into the material. This energy is set up in the 
material as an increase in the internal potential 
energy. It is "stored" energy which is 
accounted for by the increase in the mean dis- 
tance between the molecules. The source of 
this energy is the heat energy supplied. 

It is important to understand that in this 
instance the energy flowing into the material 



12 PRINCIPLES OF REFRIGERATION 



has no effect on molecular velocity (internal 
kinetic energy); only the degree of molecular 
separation (the internal potential energy) is 
affected. 

2-8. The Solid State. A material in the solid 
state has a relatively small amount of internal 
potential energy. The molecules of the material 
are rather closely bound together by each other's 
attractive forces and by the force of gravity. 
Hence, a material in the solid state has a rather 
rigid molecular structure in which the position 
of each molecule is more or less fixed and the 
motion of the molecules is limited to a vibratory 
type of movement which, depending upon the 
amount of internal kinetic energy the molecules 
possess, may be either slow or rapid. 

Because of its rigid molecular structure, a 
solid tends to retain both its size and its shape. 
A solid is not compressible and will offer 
considerable resistance to any effort to change 
its shape. 

M. The Liquid State. The molecules of a 
material in the liquid state have more energy 
than those of a material in the solid state and 
they are not so closely bound together. Their 
greater energy allows them to overcome each 
other's attractive forces to some extent and to 
have more freedom to move about. They are 
free to move over and about one another in 
such a way that the material is said to "flow." 
Although a liquid is noncompressible and will 
retain its size, because of its fluid molecular 
structure, it will not retain its shape, but 
will assume the shape of any containing 
vessel. 

2-10. The Vapor or Gaseous State. The 
molecules of a material in the gaseous state have 
an even greater amount of energy than those 
of a material in the liquid state. They have 
sufficient energy to overcome all restraining 
forces. They are no longer bound by each 
other's attractive forces, neither are they bound 
by the force of gravity. Consequently, they 
fly about at high velocities, continually collid- 
ing with each other and with the walls of the 
container. For this reason, a gas will retain 
neither its size nor its shape. It is readily com- 
pressible and will completely fill any container 
regardless of size. Further, if the gas is not 
stored in a sealed container, it will escape from 
the container and be diffused into the surround- 
ing air. 



2-11. Temperature. Temperature is a prop- 
erty of matter. It is a measure of the level of 
heat intensity or the thermal pressure of a body. 
A high temperature indicates a high level of heat 
intensity or thermal pressure, and the body is 
said to be hot. Likewise, a low temperature 
indicates a low level of heat intensity or thermal 
pressure and the body is said to be cold. 
2-12. Thermometers. The most frequently 
used instrument for measuring temperature is 
the thermometer. The operation of most ther- 
mometers depends upon the property of a liquid 
to expand or contract as its temperature is 
increased or decreased, respectively. Because 
of their low freezing temperatures and relatively 
constant coefficients of expansion, alcohol and 
mercury are the liquids most frequently used in 
thermometers. The mercury thermometer is the 
more accurate of the two because its coefficient 
of expansion is more constant through a greater 
temperature range than is that of alcohol. 
However, mercury thermometers have the dis- 
advantage of being more expensive and more 
difficult to read. Alcohol is cheaper and can be 
colored for easy visibility. 

Two temperature scales are in common use 
today. The Fahrenheit scale is used in English 
speaking countries, whereas the Centigrade 
scale is widely used in European countries as 
well as for scientific purposes. 
2-13. Centigrade Scale. The point at which 
water freezes under atmospheric pressure is 
taken as the arbitrary zero point on the Centi- 
grade scale, and the point at which water boils 
is designated as 100. The distance on the scale 
between these two points is divided into one 
hundred equal units called degrees, so that the 
distance between the freezing and boiling points 
of water on the Centigrade scale is 100°. Water 
freezes at 0° Centigrade and boils at 100° Centi- 
grade. 

2-14. Fahrenheit Scale. Although there is 
some disagreement as to the actual method used 
by Fahrenheit in designing the first temperature 
scale, it was arrived at by means similar to those 
described in the previous section. On the 
Fahrenheit scale, the point at which water 
freezes is marked as 32, and the point at which 
water boils 212. Thus, there are 180 units 
between the freezing and boiling points of 
water. The zero or reference point on the 
Fahrenheit scale is placed 32 units or degrees 



MATTER, INTERNAL ENERGY, HEAT, TEMPERATURE 13 



below the freezing point of water and is assumed 
to represent the lowest temperature Fahrenheit 
could achieve with a mixture of ammonium 
chloride and snow. 

2-15. Temperature Conversion. Temper- 
ature readings on one scale can be converted to 
reading on the other scale by using the appro- 
priate of the following equations: 

° F = 9/5° C + 32 (2-2) 

° C = 5/9(° F - 32) (2-3) 

It should be noted that the difference between 
the freezing and boiling points of water on the 
Fahrenheit scale is 180°, whereas the difference 
between these two points on the Centigrade 
scale is only 100°. Therefore, 100 Centigrade 
degrees are equivalent to 180 Fahrenheit 
degrees. This establishes a relationship such 
that 1° C equals 9/5° F (1.8° F) and 1° F equals 
5/9° C (0.555° C). This is shown graphically 
in Fig. 2-1. Since 0° on the Fahrenheit scale is 
32° F below the freezing point of water, it is 
necessary to add 32° F to the Fahrenheit equiva- 
lent after converting from Centigrade. Like- 
wise, it is necessary to subtract 32° F from a 
Fahrenheit reading before converting to Centi- 
grade. 

Example 2-1. Convert a temperature read- 
ing of 50° C to the equivalent Fahrenheit 
temperature. 

Solution. Applying 
Equation 2-2, ° F 



212* 



9/5(50° Q + 32 
= 122° F 



Example 2-2. A thermometer on the wall 
of a room reads 86° F. What is the room 
temperature in degrees Centigrade? 

Solution. Applying Formula 
2-3, the room temperature 5/9(86-32) 

in ° C = 30° C 

Example 2-3. A thermometer indicates 
that the temperature of a certain quantity of 
water is increased 45° F by the addition of 
heat. Compute the temperature rise in Centi- 
grade degrees. 

Solution. Temperature rise in ° F 

= 45°F 
Temperature rise in ° C 5/9(45° F) 

= 25°C 

2-16. Absolute Temperature. Tempera- 
ture readings taken from either the Fahrenheit 



32* 
0* 

-40" 



-460* 



Boiling point of water 



Freezing point of water 



Scales coincide 



Absolute zero 



100* 



0* 
-17.8* 

-40* 



-273' 



Fig. 2-1. Comparison of Fahrenheit and Centigrade 
temperature scales. 



or Centigrade scales are in respect to arbitrarily 
selected zero points which, as has been shown, 
are not even die same for the two scales. When 
it is desired to know only the change in tem- 
perature that occurs during a process or the 
temperature of a substance in relation to some 
known reference point, such readings are 
entirely adequate. However, when temperature 
readings are to be applied in equations dealing 
with certain fundamental laws, it is necessary 
to use temperature readings whose reference 
point is the true or absolute zero of tempera- 
ture. Experiment has indicated that such a 
point, known as Absolute Zero, exists at approxi- 
mately -460° F or -273° C (Fig. 2-1). 

Temperature readings in reference to Abso- 
lute Zero are designated as absolute tempera- 
tures and may be in either Fahrenheit or 
Centigrade degrees. A temperature reading on 
the Fahrenheit scale can be converted to 
absolute temperature by adding 460° to the 
Fahrenheit reading. The resulting temperature 
is in degrees Rankine (° R). 

Likewise, Centigrade temperatures can be 
converted to absolute temperatures by adding 
273° to the Centigrade reading. The resulting 
temperature is stated in degrees Kelvin (° K). 



14 PRINCIPLES OF REFRIGERATION 



In converting to and from absolute tempera- 
tures, the following equations will apply: 

(2-4) 



T= t+460 
t = r-460 
T « / + 273 

/ - T - 273 



(2-5) 
(2-6) 
(2-7) 



where T = absolute temperature in degrees 

Rankine or Kelvin 
t — temperature in degrees Fahrenheit 

or Centigrade 
Equations 2-4 and 2-5 apply to the Rankine and 
Fahrenheit scales, whereas Equations 2-6 and 
2-7 apply to the Kelvin and Centigrade scales. 
Hereafter in this book Rankine and Fahrenheit 
temperatures are used unless otherwise specified. 

Example 2-4. A thermometer on the tank 
of an air compressor indicates that the tempera- 
ture of the air in the tank is 95° F. Determine 
the absolute temperature in degrees Rankine. 

Solution. Applying 
Equation 2-4, T - 95° F + 460° 

= 555° R 

Example 2-5. The temperature of the 
vapor entering the suction of a refrigeration 
compressor is -20° F. Compute the tempera- 
ture of the vapor in degrees Rankine. 

Solution. Applying 
Equation 2-4, T = -20° F + 460° 

= 440°R 

Example 24. If the temperature of a gas 
is 100° C, what is its temperature in degrees 
Kelvin? 



Solution. Applying 
Equation 2-6, 



T = 100° C + 273° 
= 373° K 



Example 2-7. The temperature of steam 
leaving a boiler is 610° R. What is the tempera- 
ture of the steam on the Fahrenheit scale? 

Solution. Applying 
Equation 2-5, / = 610° R - 460° 

= 150° F 

2-17. Direction and Rate of Heat Flow. 

Heat will flow from one body to another when, 
and only when, a difference in temperature 
exists between the two bodies. If the tempera- 



ture of the two bodies is the same, there is no 
transfer of heat. 

Heat always flows down the temperature 
scale from a high temperature to a low tem- 
perature, from a hot body to a cold body, and 
never in the opposite direction. Since heat is 
energy and cannot be destroyed, if heat is to 
leave one body of material, it must flow into 
and be absorbed by another body of material 
whose temperature is below that of the body 
being cooled. 

The rate of heat transfer between two bodies 
is always directly proportional to the difference 
in temperature between the two bodies. 
2-18. Methods of Heat Transfer. The trans- 
fer of heat from one place to another occurs 
in three ways: (1) conduction, (2) convection, 
and (3) radiation. 

2-19. Conduction. Heat transfer by con- 
duction occurs when energy is transmitted by 
direct contact between the molecules of a single 
body or between the molecules of two or more 
bodies in good thermal contact with each other. 
In either case, the heated molecules communi- 
cate their energy to the other molecules im- 
mediately adjacent to them. The transfer of 
energy from molecule to molecule by conduction 
is similar to that which takes place between the 
balls on a billiard table, wherein all or some 
part of the energy of motion of one ball is trans- 
mitted at the moment of impact to the other 
balls that are struck. 

When one end of a metal rod is heated over a 
flame, some of the heat energy from the heated 
end of the rod will flow by conduction from 
molecule to molecule through the rod to the 
cooler end. As the molecules at the heated end 
of the rod absorb energy from the flame, their 
energy increases and they move faster and 
through a greater distance. The increased 
energy of the heated molecules causes them to 
strike against the molecules immediately ad- 
jacent to them. At the time of impact and 
because of it, the faster moving molecules trans- 
mit some of their energy to their slower moving 
neighbors so that they too begin to move more 
rapidly. In this manner, energy passes from 
molecule to molecule from the heated end of the 
rod to the cooler end. However, in no case 
would it be possible for the molecules furthest 
from the heat source to have more energy than 
those at the heated end. 



MATTER, INTERNAL ENERGY, HEAT, TEMPERATURE IS 



As heat passes through the metal rod, the air 
immediately surrounding the rod is also heated 
by conduction. The rapidly vibrating particles 
of the heated rod strike against the molecules 
of the air which are in contact with the rod. 
The energy so imparted to the air molecules 
causes them to move about at a higher rate and 
communicate their energy to other nearby air 
molecules. Thus, some of the heat supplied to 
the metal rod is conducted to and carried away 
by the surrounding air. 

If the heat supply to the rod is interrupted, 
heat will continue to be carried away from the 
rod by the air surrounding until the tempera- 
ture of the rod drops to that of the air. When 
this occurs, there will be no temperature differ- 
ential, the system will be in equilibrium, and 
no heat will be transferred. 

The rate of heat transfer by conduction, as 
previously stated, is in direct proportion to the 
difference in temperature between the high and 
low temperature parts. However, all materials 
do not conduct heat at the same rate. Some 
materials, such as metals, conduct heat very 
readily, whereas others, such as glass, wood, and 
cork, offer considerable resistance to the con- 
duction of heat. Therefore, for any given tem- 
perature difference, the rate of heat flow by 
conduction through different materials of the 
same length and cross section will vary with the 
particular ability of the various materials to 
conduct heat. The relative capacity of a material 
to conduct heat is known as its conductivity. 
Materials which are good conductors of heat 
have a high conductivity, whereas materials 
which are poor conductors have a low con- 
ductivity and are used as heat insulators. 

In general, solids are better conductors of 
heat than liquids, and liquids are better con- 
ductors than gases. This is accounted for by 
the difference in the molecular structure. Since 
the molecules of a gas are widely separated, the 
transfer of heat by conduction, that is, from 
molecule to molecule, is difficult. 
2-20. Convection. Heat transfer by con- 
vection occurs when heat moves from one place 
to another by means of currents which are set 
up within some fluid medium. These currents 
are known as convection currents and result 
from the change in density which is brought 
about by the expansion of the heated portion of 
the fluid. 



Cooler portions of water descend to 
replace the lighter portions that rise 



gas^fgje^* - ^^ 






Flame 




Heat is conducted 
from flame to 
water through 
bottom of vessel 



Heated portions of water become 
lighter and rise toward surface, 
thereby distributing the heat 
throughout the entire mass 

Fig. 2-2. Convection currents set up in a vessel of 
water when the vessel is heated at bottom center. 

When any portion of a fluid is heated, it ex- 
pands and its volume per unit of weight 
increases. Thus, the heated portion becomes 
lighter, rises to the top, and is immediately 
replaced by a cooler, heavier portion of the 
fluid. For example, assume that a tank of 
water is heated on the bottom at the center 
(Fig. 2-2). The heat from the flame is conducted 
through the metal bottom of the tank to the 
water inside. As the water adjacent to the heat 
source absorbs heat, its temperature increases 
and it expands. The heated portion of the 
water, being lighter than the water surrounding, 
rises to the top and is replaced by cooler, more 
dense water pushing in from the sides. As this 
new portion of water becomes heated, it too 
rises to the top and is replaced by cooler water 
from the sides. As this sequence continues, the 
heat is distributed throughout the entire mass 
of the water by means of the convection currents 
established within the mass. 

Warm air currents, such as those which occur 
over stoves and other hot bodies, are familiar to 
everyone. How convection currents are utilized 
to carry heat to all parts of a heated space is 
illustrated in Fig. 2-3. 

2-21. Radiation. Heat transfer by radiation 
occurs in the form of a wave motion similar to 
light waves wherein the energy is transmitted 
from one body to another without the need for 



16 PRINCIPLES OF REFRIGERATION 




Steam coils' 
Fig. 2-3. Room heated by natural convection. 

intervening matter. Heat energy transmitted 
by wave motion is called radiant energy. 

It is assumed that the molecules of a body are 
in rapid vibration and that this vibration sets 
up a wave motion in the ether surrounding the 
body.* Thus, the internal molecular energy of 
the body is converted into radiant energy waves. 
When these energy waves are intercepted by 
another body of matter, they are absorbed by 
that body and are converted into its internal 
molecular energy. 

The earth receives heat from the sun by 
radiation. The energy of the sun's molecular 
vibration is imparted in the form of radiant 
energy waves to the ether of interstellar space sur- 
rounding the sun. The energy waves travel across 
billions of miles of space and impress their 
energy upon the earth and upon any other 
material bodies which intercept their path. The 
radiant energy is absorbed and transformed 
into internal molecular energy, so that the 
vibratory motion of the hot body (the sun) is 
reproduced in the cooler body (the earth). 

All materials give off and absorb heat in the 
form of radiant energy. Any time the tempera- 
ture of a body is greater than that of its sur- 
roundings, it will give off more heat by radiation 
than it absorbs. Therefore, it loses energy to 
its surroundings and its internal energy de- 
creases. If the temperature of the body is below 
that of its surroundings, it absorbs more radiant 
energy than it loses and its internal energy 
increases. When no temperature difference 

* Ether is the name given to that which fills all 
space unoccupied by matter, such as interstellar 
space and the space between the molecules of every 
material. 



exists, the energy exchange is in equilibrium and 
the body neither gains nor loses energy. 

Heat transfer through a vacuum is impossible 
by either conduction or convection, since these 
processes by their very nature require that 
matter be the transmitting media. Radiant 
energy, on the other hand, is not dependent 
upon matter as a medium of transfer and there- 
fore can be transmitted through a vacuum. 
Furthermore, when radiant energy is transferred 
from a hot body to a cold body through some 
intervening media such as air the temperature of 
the intervening media is unaffected by the 
passage of the radiant energy. For example, 
heat is radiated from a "warm" wall to a "cold" 
wall through the intervening air without having 
any appreciable effect upon the temperature of 
the air. Since the molecules of the air are rela- 
tively few and widely separated, the waves of 
radiant energy can easily pass between them so 
that only a very small part of the radiant energy 
is intercepted and absorbed by the molecules of 
the air. By far the greater portion of the radiant 
energy impinges upon and is absorbed by the 
solid wall whose molecular structure is much 
more compact and substantial. 

Heat waves are very similar to light waves, 
differing from them only in length and frequency. 
Light waves are radiant energy waves of such 
length as to be visible to the human eye. Thus, 
light waves are visible heat waves. Whether 
heat waves are visible or invisible depends upon 
the temperature of the radiating body. For 
example, when metal is heated to a sufficiently 
high temperature, it will "glow," that is, emit 
visible heat waves (light). 

When radiant energy waves, either visible or 
invisible, strike a material body, they may be 
reflected, refracted, or absorbed by it, or they 
may pass through it to some other substance 
beyond. 

The amount of radiant energy which will pass 
through a material depends upon the degree of 
transparency. A highly transparent material, 
such as clear glass or air, will allow most of the 
radiant energy to pass through to the materials 
beyond, whereas opaque materials, such as 
wood, metal, cork, etc., cannot be penetrated 
by radiant energy waves and none will pass 
through. 

The amount of radiant energy which is either 
reflected or absorbed by a material depends 



MATTER, INTERNAL ENERGY, HEAT. TEMPERATURE 17 



upon the nature of the material's surface, that 
is, its texture and its color. Materials having a 
light-colored, highly polished surface, such as a 
mirror, reflect a maximum of radiant energy, 
whereas materials having rough, dull, dark 
surfaces will absorb the maximum amount of 
radiant energy. 

2-22. British Thermal Unit. It has already 
been established that a thermometer measures 
only the intensity of heat and not the quantity. 
However, in working with heat it is often 
necessary to determine heat quantities. Obvi- 
ously, some unit of heat measure is required. 

Heat is a form of energy, and as such is 
intangible and cannot be measured directly. 
Heat can be measured only by measuring the 
effects it has on a material, such as the change in 
temperature, state, color, size, etc. 

The most universally used unit of heat measure 
is the British thermal unit, abbreviated Btu. A 
Btu is denned as the quantity of heat required to 
change the temperature of 1 lb of water 1° F. 
This quantity of heat, if added to 1 lb of water, 
will raise the temperature of the water 1°F. 
Likewise, if 1 Btu is removed from 1 lb of water, 
the temperature of the water will be lowered 
1°F. 

The quantity of heat required to change the 
temperature of 1 lb of water 1° F is not a con- 
stant amount. It varies slightly with the tem- 
perature range at which the change occurs. For 
this reason, a Btu is more accurately defined as 
being l/180th of the quantity of heat required to 
raise the temperature of 1 lb of water from the 
freezing point (32° F) to the boiling point 
(212° F). This is identified as the "mean Btu" 
and is the exact amount of heat required to raise 
the temperature of 1 lb of water from 62 to 
63° F. If the change in temperature occurs at 
any other point on the temperature scale, the 
amount of heat involved is either more or less 
than the mean Btu, depending upon the par- 
ticular point on the temperature scale that the 
change takes place. However, the variation 
from the mean Btu is so slight that it may be 
neglected and, regardless of the temperature 
range, for all practical purposes it is sufficiently 
accurate to assume that the temperature of 1 lb 
of water is changed 1° F by the addition or 
removal of 1 Btu. 

2-23. Specific Heat. The specific heat of a 
material is the quantity of heat required to 



change the temperature of 1 pound of the 
material 1° F. For instance, the specific heat of 
aluminum is 0.226 Btu/lb/°F, whereas that of 
brass is 0.089 Btu/lb/°F. This means that 0.226 
Btu is required to raise the temperature of 1 
pound of aluminum 1° F, whereas only 0.089 
Btu is necessary to change the temperature of 
1 pound of brass 1° F. Note that by the defini- 
tion of the Btu the specific heat of water is 1 
Btu per pound per degree Fahrenheit. 

The specific heat of any material, like that of 
water, varies somewhat throughout the tem- 
perature scale. Here again, the variation is so 
slight that it is sufficiently accurate for most 
calculations to consider the specific heat to be 
a constant amount. This is not true, however, 
as the material passes through a change in 
physical state. The specific heat of a material 
in the solid state is approximately one-half that 
of the same -material in the liquid state. For 
instance, the specific heat of ice is 0.S Btu, 
whereas that of water is one. The specific heat 
values of materials in the gaseous state are dis- 
cussed in another chapter. 
2-24. Calculating Heat Quantity. The 
quantity of heat which must be added to or 
removed from any given mass of material in 
order to bring about a specified change in its 
temperature can be computed by using the 
following equation: 



Q, - MC(t t - t0 



(2-8) 



where Q, = the quantity of heat either absorbed 
or rejected by the material 
M = the weight of the material in 

pounds 
C = the specific heat of the material 
?! = the initial temperature 
t t = the final temperature 

Example 2-8. Twenty pounds of water at 
an initial temperature of 76° F are heated until 
the temperature is increased to 180° F. How 
much heat must be supplied? 



Solution. 
Applying 
Equation 2-8, 



Q, - 20 lb x 1 x (180 - 76) 
= 2080 Btu 



Example 2-9. If water weighs 8.33 lb per 
gallon, how much heat is rejected by 30 gal of 
water in cooling from 80° F to 35° F? 



18 PRINCIPLES OF REFRIGERATION 



Solution. 
Weight of 

water in 30 gal x 8.33 lb/gal 

pounds = 250 lb 

Applying Q, = 250 lb x 1 x (35 - 80) 
Equation 2-8, = 250 lb x 1 x (-45) 

= 11,250 Btu 

Note: Since the specific heat of a material is 
given in terms of Btu/lb/° F, the weight of the 
material must be determined before Equation 
2-8 can be applied. 

Where t t is less than t l7 the answer obtained 
by applying Equation 2-8 will be negative, 
indicating that heat is rejected by rather than 
absorbed by the material. In this type of 
problem, where the direction of heat flow is 
obvious, the negative sign can be ignored and 
the answer assumed to be positive. 

Example 2-10. Fifteen pounds of cast iron 
are cooled from 500° F to 250° F by being im- 
mersed in 3 gallons (25 lb) of water whose 
initial temperature is 78° F. Assuming that the 
specific heat of the cast iron is 0.101 Btu/lb/° F 
and that all of the heat given up by the cast iron 
is absorbed by the water, what is the final 
temperature of the water? 

Solution. By 
applying Equation 
2-8 to compute the 
total quantity of heat 
given up by the cast 
iron, 

By rearranging and 
applying Equation 2-8 
to determine the final 
temperature of the 
water after absorbing 
the heat given up by 
the cast iron, 

2-25. Heat Divided 



Q, = 15 lb x 0.101 
x 250 

= 378.75 Btu 

h MC 
378.75 



+ 'i 



+ 78°F 



into 



25 x 1 
= 15.15° +78 
= 93°F 
Two Kinds 



or 



Categories. It has been previously stated 
(Section 2-6) that heat has the ability to bring 
about a change in the physical state of a material 
as well as the ability to cause a change in its 
temperature. Heat is divided into two kinds or 
categories, depending upon which of these two 
effects it has on a material which either absorbs 
or rejects it. The division of heat into several 
classifications is made only to facilitate and 
simplify certain necessary calculations and does 
not stem from any difference in the nature of 
heat itself. 

2-26. Sensible Heat. When heat either ab- 
sorbed or rejected by a material causes or 



accompanies a change in the temperature of the 
material, the heat transferred is identified as 
sensible heat. The term sensible is applied to 
this particular heat because the change in tem- 
perature it causes can be detected with the sense 
of touch and can, of course, be measured with a 
thermometer. 

2-27. Latent Heat. When heat, either added 
to or rejected by a material, brings about or 
accompanies a change in the physical state of the 
material, the heat is known as latent heat. The 
name latent, a Latin word meaning hidden, is 
said to have been given to this special kind of 
heat by Dr. Joseph Black because it apparently 
disappeared into a material without having any 
effect on the temperature of the material. 

Many materials progressing up the tempera- 
ture scale will pass through two changes in the 
state of aggregation: first, from the solid to the 
liquid phase and then, as the temperature of 
the liquid is further increased to a certain level 
beyond which it cannot exist as a liquid, the 
liquid will change into the vapor state. When 
the change occurs in either direction between 
the solid and liquid phases, the heat involved is 
known as the latent heat of fusion. When the 
change occurs between the liquid and vapor 
phases, the heat involved is the latent heat of 
vaporization. 

2-28. Sensible Heat of a Solid. To obtain a 
better understanding of the concept of molecular 
energy, consider the progressive effects of heat 
as it is taken in by a material whose initial 
thermodynamic condition is such that its energy 
content is zero. Assume that a solid in an open 
container is at a temperature of — 460° F 
(Absolute Zero). Theoretically, at this tem- 
perature the molecules of the material have no 
energy and are completely at rest. 

When heat energy flows into the solid, the 
molecules Of the solid begin to move slowly and 
the temperature of the solid begins to climb. 
The more heat energy taken in by the solid, the 
faster the molecules vibrate and the warmer the 
solid becomes. The increase in molecular velo- 
city and in the temperature of the solid continues 
as more heat is absorbed, until the solid reaches 
its melting or fusion temperature. The total 
quantity of heat energy required to bring the 
temperature of the solid from the original con- 
dition of Absolute Zero to the melting or fusion 
temperature is known as the sensible heat of the 



MATTER, INTERNAL ENERGY, HEAT, TEMPERATURE 19 



solid. As previously shown, the quantity of heat 
which must be transferred in order to bring 
about a specified change in the temperature of 
any given mass of any material can be calculated 
by applying Equation 2-8. 
2-29. The Melting or Fusion Temperature. 
Upon reaching the fusion temperature, the 
molecules of the solid are moving as rapidly as 
is possible within the rigid molecular structure 
of the solid state. It is not possible to increase 
further the motion of the molecules or the tem- 
perature of the solid beyond this point without 
first overcoming partially the forces of mutual 
attraction which exists between the molecules. 
Hence, the material cannot exist in the solid 
state at any temperature above its melting or 
fusion temperature. On reaching the fusion 
temperature, any additional heat absorbed by 
the material will cause some part of the solid 
to revert to the liquid phase. 

The exact temperature at which melting or 
fusion occurs varies with the different materials 
and with the pressure. For instance, at normal 
atmospheric pressure, the fusion temperature of 
lead is approximately 600° F, whereas copper 
melts at approximately 2000° F and ice at only 
32° F. In general, the melting temperature 
decreases as the pressure increases except for 
noncrystalline solids, whose melting tempera- 
tures increase as the pressure increases. 
2-30. Latent Heat of Fusion. When heat 
is absorbed by a solid at the fusion temperature, 
the molecules of the solid utilize the energy to 
overcome partially their attraction for one 
another. They break away from one another to 
some extent and become more widely separated. 
As the molecules flow over and about one 
another, the material loses the rigidity of the 
solid state and becomes fluid. It can no longer 
support itself independently and will assume the 
shape of any containing vessel. 

The attraction which exists between the mole- 
cules of a solid is considerable and a relatively 
large quantity of energy is required to overcome 
that attraction. The quantity of heat required 
to melt one pound of a material from the solid 
phase into the liquid phase is called the latent 
heat of fusion. The latent heat of fusion, along 
with other values such as specific heat, fusion 
temperature, etc., for the different materials has 
been determined by experiment and may be 
found in various tables. 



It is important at this point to emphasize that 
the change of phase occurs in either direction at 
the fusion temperature, that is, the temperature 
at which the solid will melt into the liquid phase 
is the same as that at which the liquid will 
freeze into the solid phase. Further, the quan- 
tity of heat that must be rejected by a certain 
weight of liquid at the fusion temperature in 
order to freeze into the solid state is exactly 
equal to the amount of heat that must "be ab- 
sorbed by the same weight of the solid in melting 
into the liquid state. 

None of the heat absorbed or rejected during 
the change of phase has any effect on molecular 
velocity. Therefore the temperature of the 
material remains constant during the phase 
change, and the temperature of the resulting 
liquid or solid is the same as the fusion tem- 
perature.* 

The quantity of heat that is absorbed by a 
given weight of a solid at the fusion temperature 
in melting into the liquid phase, or, conversely, 
the quantity of heat that is rejected by a given 
weight of liquid at the fusion temperature in 
freezing or solidifying, can be determined by 
applying the following equation: 

Q L = M x h it (2-9) 

where Ql = the quantity of heat in Btu 

M = the mass or weight in pounds 
h it = the latent heat in Btu per pound 

Example 2-11. Calculate the quantity of 
heat required to melt 12 lb of ice at 32° F into 
water at 32° F. The latent heat of fusion of 
water under atmospheric pressure is 144 Btu 
per pound. 

Solution. Apply- 
ing Equation 2-9, the 
quantity of heat re- 
quired to melt 12 lb 
of ice 



12 lb x 144 Btu/lb 
= 1728 Btu 



Note. Since 12 lb of ice absorb 1728 Btu in 
melting into water, it follows that 12 lb of water 
at 32° F will reject 1728 Btu in returning to the 
solid state. 

* This applies with absolute accuracy only to 
crystalline solids. Noncrystalline solids, such as 
glass, have indefinite fusion temperatures. That 
is, the temperature will vary during the change of 
phase. However, for the purpose of calculating heat 
quantities, the temperature is assumed to remain 
constant during the phase change. 



20 PRINCIPLES OF REFRIGERATION 



Example 2-12. If SO lb of ice at 32° F 
absorb 6000 Btu, what part of the ice will be 
melted? 

Solution. By rearranging = _G 
and applying Equation 2-9, h if 

the part of the ice melted, M _ 6000 Btu 

~ 144 Btu/lb 
- 41.66 lb 

2-31. Sensible Heat of the Liquid. When a 
material passes from the solid to the liquid 
phase, the resulting liquid is at the fusion tem- 
perature. The temperature of the liquid may 
then be increased by the addition of heat. Any 
heat absorbed by a liquid after the change of 
state is set up in the liquid as an increase in the 
internal kinetic energy. Molecular velocity 
increases and the temperature of the liquid rises. 
But here again, as in the case of the solid, the 
temperature of the liquid eventually reaches a 
point beyond which it cannot be further in- 
creased. A liquid cannot exist as a liquid at any 
temperature above its vaporizing temperature 
for a given pressure and, upon reaching the 
vaporizing temperature, if additional heat is 
taken in by the liquid, some part of the liquid 
will change to the vapor phase. 

The total quantity of heat taken in by a liquid 
as its temperature is increased from the fusion 
to the vaporizing temperature is called the 
sensible heat of the liquid. Here again, Equation 
2-8, sometimes known as the "sensible heat 
equation," can be applied to determine the 
quantity of heat necessary to change the tem- 
perature of any given weight of liquid through 
any specified temperature range. 
2-32. Saturation Temperature. The tem- 
perature at which a liquid will change into the 
vapor phase is called the saturation temperature, 
sometimes referred to as the "boiling point" 
or "boiling temperature." A liquid whose tem- 
perature has been raised to the saturation tem- 
perature is called a saturated liquid. 

The saturation temperature, that is, the tern' 
perature at which vaporization occurs, is 
different for each liquid. Iron, for example, 
vaporizes at 4450° F, copper at 4250° F, and 
lead 3000° F. Water, of course, boils at 212° F, 
and alcohol at 170° F. Some liquids boil at 
extremely low temperatures. A few of these 
are ammonia, oxygen, and helium, which boil 
at temperatures of -28° F, -295° F, and 
—452° F, respectively. 



2-33. Latent Heat of Vaporization. Any 

heat taken in by a liquid after the liquid reaches 
the saturation temperature is utilized to increase 
the degree of molecular separation (increases the 
internal potential energy) and the substance 
changes from the liquid to the vapor phase.* 
There is no increase in molecular velocity and, 
therefore, no change in the -internal kinetic 
energy during the change in phase. Hence, 
the temperature remains constant during the 
phase change and the vapor which results is at 
the vaporizing temperature. 

As the material changes state from a liquid 
to a vapor, the molecules of the material acquire 
sufficient energy to overcome all restraining 
forces, including the force of gravity. The 
amount of energy required to do the internal 
work necessary to overcome these restraining 
forces is very great. For this reason, the capacity 
of a material to absorb heat while undergoing 
a change from the liquid to the vapor phase is 
enormous, many times greater even than its 
capacity to absorb heat in changing from the 
solid to the liquid phase. 

The quantity of heat which 1 lb of a liquid 
absorbs while changing into the vapor state is 
known as the latent heat of vaporization. The 
latent heat of vaporization, like the saturation 
temperature, is different for each material. It 
will be shown later that both the latent heat 
value and the saturation temperature of any 
particular liquid vary with the pressure over the 
liquid. When the pressure increases, the satur- 
ation temperature increases and the latent heat 
value decreases. 

The quantity of heat required to vaporize any 
given weight of liquid at the saturation tem- 
perature is calculated by the following equation: 



Q L =M xh 



fa 



(2-10) 



where Ql = the quantity of heat in Btu 
M = the mass or weight -in pounds 
h fg = the latent heat of vaporization in 
Btu/lb 

Example 2-13. If the latent heat of vapori- 
zation of water is 970 Btu per pound, how 

* Some of the energy added to the material leaves 
the material as external work and has no effect on 
the internal energy of the material. When the 
pressure is constant, the amount of external work 
done is proportional to the change in volume. 
External work is discussed in detail later. 



MATTER, INTERNAL ENERGY, HEAT, TEMPERATURE 21 



much heat is requited to vaporize 3 gal of water 
at the saturation temperature of 212° F? 



Solution. Total 
weight of water M 

Applying Equation 
2-10, Q L 



3 gal x 8.33 lb/gal 
= 25 lb 

25 lb x 970 Btu/lb 
= 24,250 Btu 



Example 2-14. One gallon of water at 
200° F in an open container absorbs 1200 Btu. 
How much water is vaporized? 

Solution. Since the saturation temperature 
of water at atmospheric pressure is 212° F, the 
entire mass of the water must be raised to this 
temperature before any water will vaporize 



Weight of 1 gal of water 

Applying Equation 2-8, 
the heat required to raise 
the temperature of the 
water from 200° F to 
212° F, 0, 

Heat available to vapor- 
ize some portion of the 
water 

Rearranging and apply- 
ing Equation 2-10, the 
weight of water vaporized, 
M 



= 8.33 lb 



8.33 lb x 
x 
= 100 Btu 



1 

12° 



1200 - 100 
1100 Btu 

1100 
970 

1.135 lb or 
0.136 gal 



Example 2-15. If 5000 Btu are removed 
from 8 lb of saturated steam at atmospheric 
pressure, how much of the steam will condense 
into water? 



Solution. By rearrang- 
ing and applying Equation 
2-10, 



M = 



5000 Btu 
970 Btu/lb 
= 5.15 lb 



2-34. Superheat — the Sensible Heat of a 
Vapor. Once a liquid has been vaporized, 
the temperature of the resulting vapor can be 
further increased by the addition of heat. The 
heat added to a vapor after vaporization is the 
sensible heat of the vapor, more commonly 
called superheat. When the temperature of a 
vapor has been so increased above the satur- 
ation temperature, the vapor is said to be 
superheated and is called a superheated vapor. 
Superheated vapors are discussed at length in 
another chapter. 
2-35. Total Heat. The total heat of a 



material at any particular condition is the sum 
total of all the sensible and latent heat required 
to bring it to that condition from an initial 
condition of Absolute Zero.* 

Example 2-16. Compute the total heat 
content of 1 lb of steam at 212° F. 

Solution. The total heat of 1 lb of saturated 
steam is the sum of the following heat quantities : 
(a) To raise the temperature of 1 lb of ice from 
-460° F to 32° F, 

= 1 x 0.5 x 

[32 - (-460)] 
= 1 x 0.5 x 492 
= 246 Btu 

(6) To melt 1 lb of ice at 32° F into water at 
32° F, 

applying Equation 2-9, Q L = 1 x 144 

= 144 Btu 

(c) To increase temperature of water from 32° F 
to 212° F, 



applying Equation 
2-8, 



applying Equation 
2-8, 



= 1 x 1 x 

(212 - 32) 
= 1 x 1 x 180 
= 180 Btu 

Q L = 1 x 970 
= 970 Btu 



(d) To vaporize 1 lb of water, 
applying Equation 2-10, 

(e) Summation: 
Sensible heat of the solid = 246 Btu 
Latent heat of fusion = 144 Btu 
Sensible heat of the liquid = 180 Btu 
Latent heat of vaporization = 970 Btu 

Total heat of 1 lb 

of steam = 1540 Btu 

Through the use of a temperature-heat 
diagram, the solution to Example 2-15 is shown 
graphically in Fig. 2-4. 

2-36. Mechanical Energy Equivalent. 

Normally the external energy of a body is ex- 
pressed in mechanical energy units (work), 
whereas the internal energy of a body is ex- 
pressed in heat energy units. 

The fact that internal energy is usually ex- 
pressed in heat energy units gives rise to the 
definition of heat as molecular or internal energy. 
As previously stated, from a thermodynamic 

* The total heat of a material is commonly known 
as "enthalpy," and is computed from some 
arbitrarily selected zero point rather than from 
Absolute Zero. See Section 4-18. 



22 PRINCIPLES OF REFRIGERATION 



300 



212 
200 



100 

32 




-400 
-460 



- .8 



Ss" m 
■= 8. 



J'S E 



#I~ I .3- 3i£Z 






c> 






i 

M * 

3~ 



F 



A 



J 1 



Constant temperature 



C/5 

r 7 



Constant temperature 



J L 



I 



J L 



_L 



J L 



J_ 



100 200 300 400 500 600 700 800 900 1000 1100 1200 1300 1400 1500 1600 1700 
Heat content (Btu) 

Fig. 2-4. Graphical analysis of the relationship of heat content to the temperature and state of a 
material. 



point of view energy is heat energy only when 
it is in transition from one body to another 
because of a difference in temperature between 
the two bodies. Once the energy flows into a 
body it becomes "stored" thermal energy. 
Hence, thermodynamically speaking, internal 
energy is not heat but thermal energy in storage. 

Not all the heat energy flowing into a body is 
stored in the body as internal energy. In many 
instances, some or all of the energy flowing into 
the body passes through or leaves the body as 
work (mechanical energy). This is made clear 
in another section. 

Furthermore, up to this point it has been 
assumed that the internal energy of a body is 
increased only by the addition of heat energy 
directly, as from a flame or some other heat 
source. However, this is not the case. The 
internal or molecular energy of a body may also 
be increased when work is done on the body. 
That is, the mechanical energy of the work done 
on a body may be converted to the internal 
energy of the body. For example, the head of a 
nail struck by a hammer will become warm as a 



part of the mechanical energy of the hammer 
blow is converted to the internal kinetic energy 
of the nail head. As the molecules of the metal 
that make up the nail head are jarred and agi- 
tated by the blow of the hammer, their motion 
or velocity is increased and the temperature of 
the nail head increases. If a wire is bent rapidly 
back and forth, the bent portion of the wire 
becomes hot because of the agitation of the 
molecules. Also, everyone is familiar with the 
increase in temperature which is brought about 
by the friction of two surfaces rubbing together. 

Often the external energy of a body is con- 
verted to internal energy and vice versa. For 
example, a bullet speeding toward a target has 
kinetic energy because of its mass and velocity. 
At the time of impact with the target, the bullet 
loses its velocity and a part of its kinetic energy 
is imparted to the molecules of both the bullet 
and the target so that the internal energy of 
each is increased. 

Since heat energy is often converted into 
mechanical energy (work) and vice versa, and 
since it is often desirable to express both the 



MATTER, INTERNAL ENERGY, HEAT, TEMPERATURE 23 



internal and external energies of a body in 
terms of the same energy unit, a factor which 
can be used to convert from one energy unit 
to the other is useful. 

It has been determined by experiment that 
one Btu of heat energy is equivalent to 778 
ft-lb of mechanical energy, that is, one Btu is the 
amount of heat energy required to do 778 ft-lb 
of work. This quantity is known as the mecha- 
nical energy equivalent and is usually repre- 
sented in equations by the symbol J. 

To convert energy in Btu into energy in 

foot-pounds, the energy in Btu is multiplied by 

778 and, to convert energy in foot-pounds to 

energy in Btu, the energy in foot-pounds is 

divided by 778. Expressed as equations, these 

relationships become 

W 
Q - -7 (2-11) 



and 



W = Q xJ 



(2-12) 



where Q = the quantity of heat energy in Btu 
W = mechanical energy or work in foot- 
pounds 
J = the mechanical energy equivalent 
of heat 

Example 2-17. Convert 36,000 ft-lb of 
mechanical energy into heat energy units. 

36,000 
Solution. Applying Equa- Q = 

tion2-ll, 



778 
= 46.3 Btu 



Example 2-18. Express 12 Btu of heat 
energy as work in mechanical energy units. 



Solution. Applying 
Equation 2-12, 



W= 12 x 778 
= 9336 ft-lb 



PROBLEMS 

1. A Fahrenheit thermometer reads 85°. What 
is the temperature in degrees Centigrade? 

Ans. 29.44° C 

2. Convert 90° Centigrade to degrees Fahren- 
heit. Ans. 194° F 

3. The temperature of a gas is 40° F. What is 
its temperature on the Rankine scale? 

Ans. 500° R 

4. The temperature of the suction vapor enter- 
ing a refrigeration compressor is —20° F. What 
is the temperature of the vapor in degrees 
Rankine? Ans. 440° R 

5. Thirty gallons of water are heated from 75° F 
to 180° F. Determine the quantity of heat 
required? Ans. 26,240 Btu 

6. In a certain industrial process, 5000 gal of 
water are cooled from 90° F to 55° F each hour. 
Determine the quantity of heat which must be 
removed each hour to produce the required 
cooling. Ans. 1,457,750 Btu 

7. Calculate the quantity of heat which must 
be removed from 60 gal of water in order to 
cool the water from 42° F and freeze it into 
ice at 32° F. Ans. 77,000 Btu 

8. If 12,120 Btu are added to 3 gal of water at 
200° F, what fraction of the water in pounds 
will be vaporized? Ans. 9.4 lb or 1.13 gal 

9. Twenty-five pounds of ice are placed in 
10 gal of water and allowed to melt. Assuming 
that there is no loss of heat to the surroundings, 
if the initial temperature of the water is 80° F, 
to what temperature will the water be cooled 
by the melting of the ice ? Ans. 35.7° 

10. A gas expanding in a cylinder does 25,000 
ft-lb of work on the piston. Determine the 
quantity of heat required to do the work. 

Ans. 32.13 Btu 



3 

Thermodynamic 
Processes 



3-1. The Effects of Heat on Volume. When 
either the velocity of the molecules or the degree 
of molecular separation is increased by the 
addition of heat, the mean distance between the 
molecules is increased and the material expands 
so that a unit weight of the material occupies a 
greater volume. This effect is in strict accord- 
ance with the theory of increased or decreased 
molecular activity as described earlier. Hence, 
when heat is added to or removed from an 
unconfined material in any of the three physical 
states, it will expand or contract, respectively. 
That is, its volume will increase or decrease with 
the addition or removal of heat. 

One of the few exceptions to this rule is water. 
If water is cooled, its volume will decrease 
normally until the temperature of the water 
drops to 39.2° F. At this point, water attains 
its maximum density and, if further cooled, its 
volume will again increase. Furthermore, after 
being cooled to 32° F, it will solidify and the 
solidification will be accompanied by still 
further expansion. In fact, 1 cu ft of water will 
freeze into approximately 1.085 cu ft of ice. 
This accounts for the tremendous expansive 
force created during solidification which is 
sufficient to burst steel pipes or other restraining 
vessels. 

The peculiar behavior of water as it solidifies 
appears to contradict the general laws govern- 
ing molecular activity as described previously. 
However, this is not the case. The unusual 
behavior of water is explained by the hypothesis 



that, although the molecules of water are 
actually closer together in the solid state than 
they are in the liquid state, they are grouped 
together to form crystals. It is the relatively 
large spaces between the crystals of the solid, 
rather than any increase in the mean distance 
between the molecules, which accounts for the 
unusual increase in volume during solidification. 
This is true also for crystalline solids other than 
ice. 

3-2. Expansion of Solids and Liquids. When 
a solid or a liquid is heated so that its tempera- 
ture is increased, it will expand a given amount 
for each degree of temperature rise. As stated 
earlier, many temperature measuring devices 
are based upon this principle. The amount of 
expansion which a material experiences with 
each degree of temperature rise is known as its 
coefficient of expansion. The coefficient of 
expansion is different for every material, and 
moreover it will vary for any particular material 
depending upon the temperature range in 
which the change occurs. 

Since solids and liquids are not readily com- 
pressible, if a solid or a liquid is restrained or 
confined so that its volume is not allowed to 
change normally with a change in temperature, 
tremendous pressures are created within the 
material itself and upon the restraining bodies, 
which is likely to cause buckling or rupturing 
of either the material, the restraining bodies, or 
both. To provide for the normal expansion and 
contraction occurring with temperature changes, 
expansion joints are built into highways, 
bridges, pipelines, etc. Likewise, liquid con- 
tainers are never completely filled. Space must 
be allowed for the normal expansion. Other- 
wise the tremendous expansive forces generated 
by a temperature increase will cause the con- 
taining vessel to rupture, sometimes with 
explosive force. 

3-3. Specific Volume. The specific volume 
of a material is the volume occupied by a 1 lb 
mass of the material. Each material has a 
different specific volume and, because of the 
change in volume which accompanies a change 
in temperature, the specific volume of every 
material varies somewhat with the temperature 
range. For instance, at 40° F, 1 lb of water has 
a specific volume of 0.01602 cu ft, whereas the 
volume occupied by 1 lb of water at 80° F is 
0.01608 cu ft. 



24 



THERMODYNAMIC PROCESSES 25 



3-4. Density. The density of a material is the 
weight in pounds of 1 cu ft of the material. 
Density is die reciprocal of specific volume, that 
is, the specific volume divided into one. The 
density of any material, like specific volume, 
varies with the temperature, but in the opposite 
direction. For example, at 40° F, the density 
of water is 62.434 lb per cubic foot (1/0.01602), 
whereas water at 80° F has a density of 62.20 
lb per cubic foot (1/0.01608). Since density and 
specific volume are reciprocals of each other, 
as one increases the other decreases. The 
density and/or the specific volume of many 
common materials can be found in various 
tables. 

The relationship between density and specific 
volume is given by the following equations: 



p= ~v 


(3-1) 


1 

V = - 

p 


(3-2) 


V = M x v 


(3-3) 


M= V x P 


(3-4) 



where v = the specific volume in cubic feet per 
pound (cu ft/lb) 
p = the density in pounds per cubic foot 

(lb/cu ft) 
V = the total volume in cubic feet 
M = the total weight in pounds 

Example 3-1 . If the specific volume of dry 
saturated steam at 212° F is 26.80 cuft per 
pound, what is the density of the steam? 



Solution. Applying 
Equation 3-1, the density p 



26.80 
= 0.0373 lb/cu ft 



Example 3-2. The basin of a cooling tower, 
measuring 5 ft x 4 ft x 1 ft, is filled with water. 
If the density of the water is 26.8 lb per cubic 
foot, what is the total weight of the water in the 
basin? 



Solution. The total volume V 5 ft x 4 ft 

x 1ft 
= 20 cu ft 
20 X 26.8 

- 536 lb 



Applying Equation 3-4, the 
total weight of water M 



3-5. Pressure-Temperature-Volume Rela- 
tionships of Gases. Because of its loose 
molecular structure, the change in the volume of 
a gas as the gas is heated or cooled is much 
greater than that which occurs in the case of a 
solid or a liquid. In the following sections, it 
will be shown that a gas may change its con- 
dition in a number of different ways and that 
certain laws have been formulated which govern 
the relationship between the pressure, temper- 
ature, and volume of the gas during these 
changes. It should be noted at the outset that 
in applying the fundamental gas laws it is 
always necessary to use absolute pressures and 
absolute temperatures in degrees Rankine. 
Further, in studying the following sections, it 
should be remembered that a gas always 
completely fills any container. 

The relationship between the pressure, 
temperature, and volume of a gas is more 
easily understood when considered through a 
series of processes in which the gas passes 
from some initial condition to some final con- 
dition in such a way that only two of these 
properties vary during any one process, whereas 
the third property remains unchanged or con- 
stant. 

3-6. Temperature-Volume Relationship at 
a Constant Pressure. If a gas is heated 
under such conditions that its pressure is kept 
constant, its volume will increase 1/492 of its 
volume at 32° F for each 1° F increase in its 
temperature. Likewise, if a gas is cooled at a 
constant pressure, its volume will decrease 1/492 
of its volume at 32° F for each 1 ° F decrease in 
its temperature. 

In order to better visualize a constant pressure 
change in condition, assume that a gas is con- 
fined in a cylinder equipped with a perfectly 
fitting, frictionless piston (Fig. 3-la). The 
pressure of the gas is that which is exerted on the 
gas by the weight of the piston and by the weight 
of the atmosphere on top of the piston. Since 
the piston is free to move up or down in the 
cylinder, the gas is allowed to expand or con- 
tract, that is, to change its volume in such a way 
that the pressure of the gas remains constant. 
As the gas is heated, its temperature and volume 
increase and the piston moves- upward in the 
cylinder. As the gas is cooled, its temperature 
and volume decrease and the piston moves down- 
ward in the cylinder. In either case, the pressure 



26 PRINCIPLES OF REFRIGERATION 



Perfectly fitting 

frictionless 

piston 



: Piston 



■■?< P = 100 psia 
£fr = 500*R J& 
&V=Icuft 



[tttl 


I 


Hi i Pistoh|| 


ft? '.■ 'Of/^'"""-"-"^ 




i£f P = 100 psia $ 


$v;T=1000*R 4. 





1 

Volu 
chai 

JL 



Volume 
change 



fTTT 

Heat 
added 




Heat 
removed 



(a) (b) (c) 

Fig. 3-1. Constant pressure process, (a) Gas confined In a cylinder with a perfectly fitting, frictionless piston, 
(b) As gas is heated, both the temperature and the volume of the gas Increase. The increase in volume is 
exactly proportional to the increase in absolute temperature, (c) As gas is cooled, both the temperature and 
the volume of the gas decrease. The decrease in volume is exactly proportional to the decrease in absolute 
temperature. 



of the gas remains the same or unchanged during 
the heating or cooling processes. 
3-7. Charles' Law for a Constant Pressure 
Process. Charles' law for a constant pressure 
process states in effect that, when the pressure 
of the gas remains constant, the volume of the 
gas varies directly with its absolute temperature. 
Thus, if the absolute temperature of a gas is 
doubled while its pressure is kept constant, its 
volume will also be doubled. Likewise, if the 
absolute temperature of a gas is reduced by 
one-half while the pressure is kept constant, its 
volume will also be reduced by one-half. This 
relationship is illustrated in Figs. 3-16 and 
3-lc. 

Charles' law for a constant pressure process 
written as an equation is as follows: When the 
pressure is kept constant, 



TrV t = T^ 



(3-5) 



where T ± = the initial temperature of the gas in 

degrees Rankine 
r 2 = the final temperature of the gas in 

degrees Rankine 
Vx = the initial volume of the gas in 

cubic feet 
V s = the final volume of the gas in cubic 

feet 



When any three of the preceding values are 
known, the fourth may be calculated by applying 
Equation 3-5. 

Example 3-3. A gas, whose initial tem- 
perature is 520° R and whose initial volume is 
5 cu ft, is allowed to expand at a constant 
pressure until its volume is 10 cu ft. Determine 
the final temperature of the gas in degrees 
Rankine. 

520 x 10 



Solution. By rearranging and 
applying Equation 3-5, the final 
temperature of the gas T 2 



= 1040° R 

Example 3-4. A gas, having an initial 
temperature of 80° F, is cooled at a constant 
pressure until its temperature is 40° F. If the 
initial volume of the gas is 8 cu ft, what is its 
final volume? 

Solution. Since the temperatures are given 
in degrees Fahrenheit, they must be converted 
to degrees Rankine before being substituted in 
Equation 3-5. 

By rearranging and apply- _ ^i^i 
ing Equation 3-5, the final T t 

volume V 2 = 500 x 8 

540 
= 7.4074 cu ft 



THERMODYNAMIC PROCESSES 27 



3-8. Pressure-Volume Relationship at a 
Constant Temperature. When the volume 
of a gas is increased or decreased under such 
conditions that the temperature of the gas does 
not change, the absolute pressure will vary 
inversely with the volume. Thus, when a gas is 
compressed (volume decreased) while its tem- 
perature remains unchanged, its absolute 
pressure will increase in proportion to the 
decrease in volume. Similarly, when a gas is 
expanded at a constant temperature, its abso- 
lute pressure will decrease in proportion to the 
increase in volume. This is a statement of 
Boyle's law for a constant temperature process 
and is illustrated in Figs. 3-2a, 3-26, and 3-2c. 
It has been previously stated that the mole- 
cules of a gas are flying about at random and at 
high velocities and that the molecules of the 
gas frequently collide with one another and with 
the walls of the container. The pressure exerted 
by the gas is a manifestation of these molecular 
collisions. Billions and billions of gas mole- 
cules, traveling at high velocities, strike the 
walls of the container during each fraction of a 
second. It is this incessant molecular bombard- 
ment which produces the pressure that a gas 
exerts upon the walls of its container. The 
magnitude of the pressure exerted depends upon 
the force and frequency of the molecular im- 
pacts upon a given area. The greater the force 



and frequency of the impacts, the greater is the 
pressure. The number of molecules confined in 
a given space and their velocity will, of course, 
determine the force and the frequency of the 
impacts. That is, the greater the number of 
molecules (the greater the quantity of gas) and 
the higher the velocity of the molecules (the 
higher the temperature of the gas), the greater 
is the pressure. The force with which the mole- 
cules strike the container walls depends only 
upon the velocity of the molecules. The higher 
the velocity the greater is the force of impact. 
The greater the number of molecules in a given 
space and the higher the velocity the more often 
the molecules will strike the walls. 

When a gas is compressed at a constant 
temperature, the velocity of the molecules re- 
mains unchanged. The increase in pressure 
which occurs is accounted for by the fact that the 
volume of the gas is diminished and a given 
number of gas molecules are confined in a smaller 
space so that the frequency of impact is greater. 
The reverse of this holds true, of course, when 
the gas is expanded at a constant temperature. 

Any thermodynamic process which occurs in 
such a way that the temperature of the working 
substance does not change during the process is 
called an isothermal (constant temperature) 
process. 

Boyle's law for a constant temperature process 



Fig. 3-2. Constant tempera- 
ture process, (a) Initial condi- 
tion, (b) Constant temperature 
expansion — volume change is 
inversely proportional to the 
change in absolute pressure. 
Heat must be added during 
expansion to keep tempera- 
ture constant, (c) Constant 
temperature compression — 
volume change is inversely 
proportional to the change in 
absolute pressure. Heat must 
be removed during com- 
pression to keep temperature 
constant. 





L 












mi: piston: mi 








!! 

-■,■ 1 I-' 
1 ■ 1 ' 

'■:■■: II ■■;'.:.• '■ 
"■^:'l l:---'v 
-__-l-L_ 


1 

i 
i 


i 
l 
i 


miipiswiiu 








<p=ibopsi | 

V=lcuft 
! T=70*F 




P * 50 psi 
. V«2cuft $ 
? T=70*F ; 








HI|liPiston:i|| 




iP m 200 psi<2 
IV = 0.5cuft1 
!§J>.70*F 


(c 


>> 




ft 

Heat 
(1 


'If 

idded 




H'l 

Heat re 
U 


ii 

moved 
) 



28 PRINCIPLES OF REFRIGERATION 



Volume = 1 cu ft 



Pressure = 100 psia 



Temperatures 500° R, 



Volume = 1 cu ft @> 



Pressure = 200 psia 



Temperature= 1000° R 



TT 



Volume = 1 cu ft 



Pressure = 50 psia 



Temperature = 250' R 



TTT 



Heat added 



Heat removed 



(a) (b) (c) 

Fig. 3-3. Constant volume process, (a) Initial condition, (b) The absolute pressure increases in direct 
proportion to the increase in absolute temperature, (c) The absolute pressure decreases in direct proportion 
to the decrease in absolute temperature. 



is represented by the following equation: if 
the temperature is constant, 

PiVi = P*r* (3-6) 

where P x = the initial absolute pressure 
P a = the final absolute pressure 
V x = the initial volume in cubic feet 
K 2 = the final volume in cubic feet 

Example 3-5. Five pounds of air are ex- 
panded at a constant temperature from an 
initial volume of 5 cu ft to a final volume of 
10 cu ft. If the initial pressure of the air is 
20 psia, what is the final pressure in psia? 

20 x 5 



initial pressure of the gas is 3000 psfa, determine 
the final pressure in psig. 



Solution. By rearranging and 
applying Equation 3-6, the final 
pressure P 2 



10 
= 10 psia 

Example 3-6. Four cubic feet of gas are 
allowed to expand at a constant temperature 
from an initial pressure of 1500 psfa to a final 
pressure of 900 psfa. Determine the final 
volume of the gas. 

_ PiVi 

P* 
_ 1500 x 4 

900 
= 6.67 cu ft 



Solution. By rearranging and 
applying Equation 3-6, the 
final volume V z 



Example 3-7. A given weight of gas, whose 
initial volume is 10 cu ft, is compressed iso- 
thermally until its volume is 4 cu ft. If the 



Solution. By rearranging 
and applying Equation 3-6, 
the final pressure P t 


PiVi 

3000 x 10 
4 
= 7500 psfa 




Dividing by 144 


7500 
144 


Subtracting the atmos- 
pheric pressure 


= 52.08 psia 
= 52.08 - 14.7 
= 37.38 psig 



3-9. Pressure-Temperature Relationship 
at a Constant Volume. Assume that a gas 
is confined in a closed cylinder so that its 
volume cannot change as it is heated or cooled 
(Fig. 3-3a). When the temperature of a gas is 
increased by the addition of heat, the absolute 
pressure will increase in direct proportion to the 
increase in absolute temperature (Fig. 3-3b). 
If the gas is cooled, the absolute pressure of the 
gas will decrease in direct proportion to the 
decrease in absolute temperature (Fig. 3-3c). 

Whenever the temperature (velocity of the 
molecules) of a gas is increased while the volume 
of the gas (space in which the molecules are 
confined) remains the same, the magnitude of 
the pressure (the force and frequency of mole- 
cular impacts on the cylinder walls) increases. 
Likewise, when a gas is cooled at a constant 
volume, the force and frequency of molecular 
impingement on the walls of the container 



diminish and the pressure of the gas will be less 
than before. The reduction in the force and the 
frequency of molecular impacts is accounted 
for by the reduction in molecular velocity. 
3-10. Charles' Law for a Constant Volume 
Process. Charles' law states in effect that 
when a gas is heated or cooled under such con- 
ditions that the volume of the gas remains 
unchanged or constant, the absolute pressure 
varies directly with the absolute temperature. 
Charles' law may be written as the following 
equation: when the volume is the same, 



THERMODYNAMIC PROCESSES 29 

is used, then V will become the specific volume 
v, and Equation 3-8 may be written: 



Vi = Vi 



(3-7) 



where T x = the initial temperature in degrees 

Rankine 
r 2 = the final temperature in degrees 

Rankine 
P x = the initial pressure in pounds per 

square inch absolute 
P 2 = the final pressure in pounds per 

square inch absolute 

Example 3-8. A certain weight of gas con- 
fined in a tank has an initial temperature of 
80° F and an initial pressure of 30 psig. If the 
gas is heated until the final gage pressure is 
50 psi, what is the final temperature in degrees 
Fahrenheit? 



Solution. By rearrang- 
ing and applying Equa- 
tion 3-7, 



Converting Rankine to 
Fahrenheit 



T, = 



T, xi». 



Pi 

(80 + 460) 
_ x(50 + 14.7) 
(30 + 14.7) 
= 782° R 

782 - 460 
= 322° F 



3-11. The General Gas Law. Combining 
Charles' and Boyle's laws produces the following 
equation: 

P -P= P -P 0-8) 

Equation 3-8 is a statement that for any given 
weight of a gas, the product of the pressure in 
psfa and the volume in cubic feet divided by the 
absolute temperature In degrees Rankine will 
always be a constant. The constant, of course, 
will be different for different gases and, for any 
one gas, will vary with the weight of gas involved. 
However, if, for any one gas, the weight of 1 lb 



Pv 
T 



(3-9) 



where R = the gas constant 



The gas constant R is different for each gas. 
The gas constant for most common gases can 
be found in tables. A few of these are given 
in Table 3-1. 

Multiplying both sides of Equation 3-9 by M 
produces 

PMv = MRT 
but since 

Mv = V 
then 

PV = MRT (3-10) 

where P = the pressure in psfa 
V = the volume in ft 3 
M = the mass in pounds 
R = the gas constant 
T = the temperature in ° R 

Equation 3-10 is known as the general gas law 
and is very useful in the solution of many 
problems involving gases. Since the value of R 
for most gases can be found in tables, if any three 
of the four properties, P, V, M, and T, are known, 
the fourth property can be determined by Equa- 
tion 3-10. Notice that the pressure must be in 
pounds per square foot absolute. 

Example 3-9. The tank of an air compres- 
sor has a volume of 5 cu ft and is filled with air 
at a temperature of 100° F. If a gage on the 
tank reads 151.1 psi, what is the weight of the 
air in the tank? 



Solution. From 
Table 3-1, R for air 

By rearranging 
and applying Equa- 
tion 3-10, the weight 
of air M 



= 53.3 

(151.1 + 14.7) 

x 144x5 

~ 53.3 x (100 + 460) 
_ 165.8 x 144 x 5 

53.3 x 560 
= 4 lb 

Example 3-10. Two pounds of air have a 
volume of 3 cu ft. If the pressure of the air is 
135.3 psig, what is the temperature of the air in 
degrees Fahrenheit? 



30 PRINCIPLES OF REFRIGERATION 



Solution. From 




Table 3-1, R for air 


= 53.3 


By rearranging 
and applying Equa- 
tion 3-10, the tem- 
perature of the air 


T- PV 

~ MR 

(135.3 + 14.7) 

x 144x3 


in °R, 


2 x 53.3 




150 x 144 x 3 




2 x 53.3 




= 607.9° R 


Converting to 
Fahrenheit, 


t = 607.9 - 460 
= 147.9° F 



3-12. External Work. Whenever a material 
undergoes a change in volume, work is done. If 
the volume of the material increases.work is done 
by the material. If the volume of the material 
decreases, work is done on the material. For 
example, consider a certain weight of gas con- 
fined in a cylinder equipped with a movable 
piston (Fig. 3-la). As the gas is heated, its 
temperature increases and it expands, moving 
the piston upward in the cylinder against the 
pressure of the atmosphere. Work is done in 
that the weight of the piston is moved through a 
distance(Fig. 3-1A).* The agency doing the work 
is the expanding gas. 

In order to do work, energy is required 
(Section 1-12). In Fig. 3-lZ>, the energy required 
to do the work is supplied to the gas as the gas 
is heated by an external source. It is possible, 
however, for a gas to do external work without 
the addition of energy from an external source. 
In such cases, the gas does the work at the 
expense of its own energy. That is, as the gas 
expands and does work, its internal kinetic energy 
(temperature) decreases in an amount equal to 
the amount of energy required to do the work. 

When a gas is compressed (its volume de- 
creased), a certain amount of work must be done 
on the gas in order to compress it. And, an 
amount of energy equal to the amount of work 
done will be imparted to the molecules of the 
gas during the compression. That is, the 
mechanical energy of the piston motion will be 
transformed into the internal kinetic energy of 
the gas (molecular motion) and, unless the gas is 
cooled during the compression, the temperature 
of the gas will increase in proportion to the 

* Some work is done, also, in overcoming friction 
and in overcoming the pressure of the atmosphere. 



amount of work done. The increase in the tem- 
perature of a gas as the gas is compressed is a 
common phenomenon and may be noted by 
feeling the valve stem of a tire being filled with a 
hand pump, or the head of an air compressor, 
etc. 

3-13. The General Energy Equation. The 
law of conservation of energy clearly indicates 
that the energy transferred to a body must be 
accounted for in its entirety. It has been shown 
that some part (or all) of the energy taken in by a 
material may leave the material as work, and 
that only that portion of the transferred energy 
which is not utilized to do external work remains 
in the body as "stored thermal energy." It is 
evident then that all of the energy transferred to 
a body must be accounted for in some one or in 
some combination of the following three ways: 

(1) as an increase in the internal kinetic energy, 

(2) as an increase in the internal potential energy, 
and (3) as external work done. The general 
energy equation is a mathematical statement of 
this concept and may be written: 



AQ = AK + AP + AW 



(3-11) 



where AQ = the heat energy transferred to the 
material in Btu 
AK = that fraction of the transferred 
energy which increases the in- 
ternal kinetic energy 
AP = that fraction of the transferred 
energy which increases the in- 
ternal potential energy 
A W = that fraction of the transferred 
energy which is utilized to do 
external work 

The Greek letter, A (delta), used in front of a 
term in an equation identifies a change of 
condition. For instance, where K represents 
the internal kinetic energy, AK represents the 
change in the internal kinetic energy. 

Depending upon the particular process or 
change in condition that the material undergoes, 
any of the terms in Equation 3-11 may have any 
value either positive or negative, or any may be 
equal to zero. This will be made clear later. 
3-14. External Work of a Solid or Liquid. 
When heat added to a material in either the 
solid or liquid state increases the temperature 
of the material, the material expands somewhat 
and a small amount of work is done. However, 



THERMODYNAMIC PROCESSES 31 



the increase in volume and the external work 
done is so slight that the portion of the trans- 
ferred energy which is utilized to do external 
work or to increase the internal potential energy 
is negligible. For all practical purposes, it can 
be assumed that all the energy added to a solid 
or a liquid during a temperature change in- 
creases the internal kinetic energy. None leaves 
the material as work and none is set up as an 
increase in the internal potential energy. In 
this instance, both AP and AW of Equation 
3-11 are equal to zero and, therefore, AQ is 
equal to AK . 

When a solid melts into the liquid phase, 
the change in volume is again so slight that the 
external work done may be neglected. Further- 
more, since the temperature also remains 
constant during the phase change, none of the 
transferred energy increases the internal kinetic 
energy. All the energy taken in by the melting 
solid is set up as an increase in the internal 
potential energy. Therefore, AK and AW are 
both equal to zero and A Q is equal to AP. 

This is not true, however, when a liquid 
changes into the vapor phase. The change in 
volume that occurs and therefore the external 
work done as the liquid changes into a vapor is 
considerable. For example, when 1 lb of water 
at atmospheric pressure changes into a vapor, 
its volume increases from 0.01671 cu ft to 26.79 
cu ft. Of the 970.4 Btu required to vaporize 1 
lb of water, approximately 72 Btu of this energy 
are required to do the work of expanding 
against the pressure of the atmosphere. The 
remainder of the energy is set up in the vapor as 
an increase in the internal potential energy. In 
this instance, only AK is equal to zero, so that 
AQ is equal to AP plus AW. 
3-15. "Ideal" or "Perfect" Gas. The various 
laws governing the pressure-volume-temperature 
relationships of gases as discussed in this chapter 
apply with absolute accuracy only to a hypo- 
thetical "ideal" or "perfect" gas. A "perfect" 
gas is described as one in such a condition that 
there is no interaction between the molecules of 
the gas. The molecules of such a gas are 
entirely free and independent of each other's 
attractive forces. Hence, none of the energy 
transferred either to or from an ideal gas has 
any effect on the internal potential energy. 

The concept of an ideal or perfect gas greatly 
simplifies the solution of problems concerning 



the changes in the condition of a gas. Many 
complex problems in mechanics are made simple 
by die assumption that no friction exists, the 
effects of friction being considered separately. 
The function of an ideal gas is the same as that 
of the frictionless surface. An ideal gas is 
assumed to undergo a change of condition 
without internal friction, that is, without the 
performance of internal work in the overcoming 
of internal molecular forces. 

The idea of internal friction is not difficult to 
comprehend. Consider that a liquid such as oil 
will not flow readily at low temperatures. This 
is because of the internal friction resulting from 
strong intermolecular forces within the liquid. 
However, as the liquid is heated and the mole- 
cules gain additional energy, the intermolecular 
forces are overcome somewhat, internal friction 
diminishes, and the liquid flows more easily. 

Vaporization of the liquid, of course, causes 
a greater separation of the molecules and 
brings about a substantial reduction in internal 
friction, but some interaction between the 
molecules of the vapor still exists. In the 
gaseous state, intermolecular forces are greatest 
when the gas is near the liquid phase and 
diminish rapidly as the gas is heated and its 
temperature rises farther and farther above the 
saturation temperature. A gas approaches the 
ideal state when it reaches a condition such that 
the interaction between the molecules and 
hence, internal friction, is negligible. 

Although no such thing as an ideal or perfect 
gas actually exists, many gases, such as air, 
nitrogen, hydrogen, helium, etc., so closely 
approach the ideal condition that any errors 
which may result from considering them to be 
ideal are of no consequence for all practical 
purposes. 

Although it is important that the student of 
refrigeration understand and be able to apply 
the laws of perfect gases, it should be under- 
stood that gases as they normally occur in the 
mechanical refrigeration cycle are close to the 
saturation curve, that is, they are vapors, and 
do not even approximately approach the 
condition of an ideal or perfect gas.* They 
follow the gas laws in only a very general way, 

* A vapor is sometimes defined as a gas at a 
condition close enough to the saturation curve so 
that it does not follow the ideal gas laws even 
approximately. 



32 PRINCIPLES OF REFRIGERATION 

and therefore the use of the gas laws to deter- 
mine the pressure-volume-temperature relation- 
ships of such vapors will result in considerable 
inaccuracy. In working with vapors, it is usually 
necessary to use values which have been deter- 
mined experimentally and are tabulated in 
saturated and superheated vapor tables. These 
tables are included as a part of this textbook 
and are discussed later. 

3-16. Processes for Ideal Gases. A gas is 
said to undergo a process when it passes from 
some initial state or condition to some final 
state or condition. A change in the condition of 
a gas may occur in an infinite number of ways, 
only five of which are of interest. These are the 
(1) constant pressure (isobaric), (2) constant 
volume (isometric), (3) constant temperature 
(isothermal), (4) adiabatic, and (5) polytropic 
processes. 

In describing an ideal gas, it has been said 
that the molecules of such a gas are so far apart 
that they have no attraction for one another, 
and that none of the energy absorbed by an 
ideal gas has any effect on the internal potential 
energy. It is evident, then, that heat absorbed 
by an ideal gas will either increase the internal 
kinetic energy (temperature) of the gas or it will 
leave the gas as external work, or both. Since 
the change in the internal potential energy, AP, 
will always be zero, the general energy equation 
for an ideal gas may be written: 



AQ = AK + AW 



(3-12) 



In order to better understand the energy 
changes which occur during the various pro- 
cesses, it should be kept in mind that a change in 
the temperature of the gas indicates a change in 
the internal kinetic energy of the gas, whereas a 
change in the volume of the gas indicates work 
done either by or on the gas. 
3-17. Constant Volume Process. When a 
gas is heated while it is so confined that its 
volume cannot change, its pressure and tempera- 
ture will vary according to Charles' law (Fig. 
3-3). Since the volume of the gas does not 
change, no external work is done and AW is 
equal to zero. Therefore, for a constant 
volume process, indicated by the subscript v, 



*Qv - Mv 



(3-13) 



transferred to the gas increases the internal 
kinetic energy of the gas. None of the energy 
leaves the gas as work. 

When a gas is cooled (heat removed) while 
its volume remains constant, all the energy 
removed is effective in reducing the internal 
kinetic energy of the gas. It should be noted 
that in Equation 3-12, AQ represents heat 
transferred to the gas, AK represents an in- 
crease in the internal kinetic energy, and AW 
represents work done by the gas. Therefore, 
if heat is given up by the gas, AQ is negative. 
Likewise, if the internal kinetic energy of the gas 
decreases, AK is negative and, if work is done on 
the gas, rather than by it, A W is negative. Hence, 
in Equation 3-13, when the gas is cooled, both 
AQ and AK are negative. 
3-18. Constant Pressure Process. If the 
temperature of a gas is increased by the addition 
of heat while the gas is allowed to expand so 
that its pressure is kept constant, the volume of 
the gas will increase in accordance with Charles' 
law (Fig. 3-1). Since the volume of the gas 
increases during the process, work is done by 
the gas at the same time that its internal energy 
is increased. Hence, while one fraction of the 
transferred energy increases the store of internal 
kinetic energy, another fraction of the trans- 
ferred energy leaves the gas as work. For a 
constant pressure process, identified by the 
subscript/!, the energy equation may be written 



AQ P =AK 1I +AW 



(3-14) 



Equation 3-13 is a statement that during a 
constant volume process all of the energy 



3-19. Specific Heat of Gases. The quantity 
of heat required to raise the temperature of 
1 lb of a gas 1° F while the volume of the gas 
remains constant is known as the specific heat 
at a constant volume (C„). Similarly, the 
quantity of heat required to raise the tempera- 
ture of 1 lb of a gas 1° F while the gas expands 
at a constant pressure is called the specific 
heat at a constant pressure (C„). For any 
particular gas, the specific heat at a constant 
pressure is always greater than the specific heat 
at a constant volume. The reason for this is 
easily explained. 

The quantity of energy required to increase 
the internal kinetic energy of a gas to the extent 
that the temperature of the gas is increased 1 ° F 
is exactly the same for all processes. Since, 
during a constant volume process, no work is 
done, the only energy required is that which 



THERMODYNAMIC PROCESSES 33 



increases the internal kinetic energy. However, 
during a constant pressure process, the gas 
expands a fixed amount for each degree of 
temperature rise and a certain amount of 
external work is done. Therefore, during a 
constant pressure process, energy to do the 
work that is done must be supplied in addition 
to that which increases the internal kinetic 
energy. For example, the specific heat of air 
at a constant volume is 0.169 Btu per pound, 
whereas the specific heat of air at a constant 
pressure is 0.2375 Btu per pound. For either 
process, the increase in the internal energy of 
the air per degree of temperature rise is 0.169 
Btu per pound. For the constant pressure pro- 
cess, the additional 0.068S Btu per pound 
(0.2375 — 0.169) is the energy required to do 
the work resulting from the volume increase 
accompanying the temperature rise. 

The specific heat of a gas may take any value 
either positive or negative, depending upon the 
amount of work that the gas does as it expands. 
3-20. The Change in Internal Kinetic 
Energy. During any process in which the 
temperature of the gas changes, there will be a 
change in the internal kinetic energy of the gas. 
Regardless of the process, when the tempera- 
ture of a given weight of gas is increased or 
decreased, the change in the internal kinetic 
energy can be determined by the equation 



AK = MC v (t t - fj 



(3-15) 



where AK = the increase in the internal kinetic 
energy in Btu 
M = the weight in pounds 
C v = constant volume specific heat 
t t = the final temperature 
f x = the initial temperature 

Note. The temperature may be in either 
Fahrenheit or Rankine, since the difference in 
temperature will be the same in either case as 
long as the units are consistent. 

Example 3-11. The temperature of 5 lb 
of air is increased by the addition of heat from 
an initial temperature of 75° F to a final tem- 
perature of 140° F. If C„ for air is 0.169 Btu, 
what is the increase in the internal energy? 

Solution. Applying =5 x 0.169 
Equation 3-15, the in- x (140 - 75) 

crease in internal kinetic = 5 x 0.169 x 65 

energy AK = 54.9 Btu 



Example 3-12. Twelve pounds of air are 
cooled from an initial temperature of 95° F to 
a final temperature of 72° F. Compute the 
increase in the internal kinetic energy. 



Solution. Apply- 
ing Equation 3-15, 
the increase in in- 
ternal kinetic energy 
AK 



= 12 x 0.169 

x (72 - 95) 
= 12 x 0.169 x (-23) 
= -46.64 Btu 



In Example 3-12, AK is negative, indicating 
that the gas is cooled and that the internal 
kinetic energy is decreased rather than increased. 
3-21. Heat Transferred during a Constant 
Volume Process. For a constant volume 
process, since 

then 



AQ V = AK V 
AQ V = MC v {ti - /,) (3-16) 



Example 3-13. If, in Example 3-11, the gas 
is heated while its volume is kept constant, 
what is the quantity of heat transferred to the 
gas during the process ? 



Solution. Apply- 
ing Equation 3-16, 
the heat transferred 
to the gas, 

Alternate Solution. 
From Example 
3-12, 

Since 
AQ V = AK V , 



AQ V = 5 x 0.169 

x (140 - 75) 
= 5 x 0.169 x 65 
= 54.9 Btu 



AK V = 54.9 Btu 
AQ V = 54.9 Btu 



Example 3-14. If, in Example 3-12, the 
air is cooled while its volume remains constant, 
what is the quantity of heat transferred to the 
air during the process? 



Solution. From 
Example 3-12, 
Since AQ„ = AK V . 



AK V = -46.67 Btu 
AQ V 46.67 Btu 



In Example 3-14, notice that since AK V is 
negative, indicating a decrease in the internal 
kinetic energy, AQ V must of necessity also be 
negative, indicating that heat is transferred 
from the gas rather than to it. 
3-22, External Work during a Constant 
Pressure Process. It will now be shown that 
the work done during a constant pressure 
process may be evaluated by the equation : 



W'PiVt-VJ 



(3-17) 



34 PRINCIPLES OF REFRIGERATION 



where W = the work done in foot-pounds 
P = the pressure in psfa 
Kg = the final volume in cubic feet 
V x = the initial volume in cubic feet 

Assume that the piston in Fig. 3-lc has an 
area of A square feet and that the pressure of 
the gas in the cylinder is P pounds per square 
foot. Then, the total force exerted on the top of 
the piston will be PA pounds, or 

F = P xA 

Assume now that the gas in the cylinder, having 
an initial volume V ly is heated and allowed to 
expand to volume V 2 while its pressure is kept 
constant. In doing so, the force PA acts through 
the distance 1 and work is done. Hence, 



but since 



then 



W = P xA x 1 



A x 1 = (V s - VJ 
W = P(V 2 -VJ 



Example 3-15. One pound of air having 
an initial volume of 13.34 cu ft and an initial 
temperature of 70° F is heated and allowed to 
expand at a constant pressure of 21 17 psfa to a 
final volume of 15 cu ft. Determine the amount 
of external work in foot-pounds. 



Solution. Applying 
Equation 3-17, the work in 
foot-pounds W 



= 2117 x 

(15 - 13.34) 
= 2117 x 1.66 
= 3514 ft-lb 



In Equation 3-12, AWS& always given in heat 
energy units. By application of the mechanical 
energy equivalent (Section 3-16), W in foot- 
pounds may be expressed as A W in Btu. The 
relationship is 

W 
AW = — (3-18) 



W = AW xJ 



(3-19) 



Example 3-16. Express the work done in 
Example 3-15 in terms of heat energy units. 

Solution. Applying Equation _ 3514 
3-18, the work in Btu W 778 

= 4.52 Btu 

3-23. Heat Transferred during a Constant 
Pressure Process. According to Equation 
3-14, AQ„ the total heat transferred to a gas 
during a constant pressure process is equal to 



the sum of AK P , the increase in internal kinetic 
energy, and AW 9 , the heat energy equivalent of 
the work of expansion. 

Example 3-17. Compute the total heat 
energy transferred to the air during the constant 
pressure process described in Example 3-15. 



Solution. Con- 






verting 70° F to 
degrees Rankine, 


° R = 70 + 460 
= 530° R 




Applying 
Charles' law, 


T — *1 X ^1 

2 " v x 




Equation 3-5, to 
determine the final 


_ 530 x 15 
13.34 




temperature, 


= 596" R 




Applying Equa- 
tion 3-15, the in- 
crease in internal 


AA: = 1 x 0.169 
x (596 - 
= 1 x 0.169 


530) 
x 66 


kinetic energy, 


= 11.154 Btu 




From Example 
3-15 and 3-16, 


DW, =4.52 Btu 





Applying Equa- AQ V = 11.15 + 4.52 
tion 3-14, = 15.67 Btu 

Since the specific heat at a constant pressure 
C P takes into account not only the increase in 
internal energy per pound but also the work 
done per pound per degree of temperature rise 
during a constant pressure expansion, for the 
constant pressure process only, AQ P may be 
determined by the following equation : 



AQ P - MC p (t 2 - tj) 



(3-20) 



Hence, an alternate solution to Example 3-17 is 

Applying Equation AQ V = 1 x 0.2375 

x (596 - 530) 
= 1 x 0.2375 

x 66 
= 15.67 Btu 



3-20, 



3-24. Pressure-Volume (PV) Diagram. 
Equation 3-8 is a statement that the thermo- 
dynamic state of a gas is adequately described 
by any two properties of the gas. Hence, using 
any two properties of the gas as mathematical 
coordinates, the thermodynamic state of a gas 
at any given instant may be shown as a point 
on a chart. Furthermore, when the conditions 
under which a gas passes from some initial state 
to some final state are known, the path that the 
process follows may be made to appear as a line 
on the chart. 



THERMODYNAMIC PROCESSES 35 



The graphical representation of a process or 
cycle is called a process diagram or a cycle 
diagram, respectively, and is a very useful tool 
in the analysis and solution of cyclic problems. 

Since work is a function of pressure and 
volume, when it is the work of a process or 
cycle which is of interest, the properties used as 
coordinates are usually the pressure and the 
volume. When the pressure and volume are 
used as coordinates to diagram a process or 
cycle, it is called a pressure-volume (PV) 
diagram. 

To illustrate the use of the PV diagram, a 
pressure-volume diagram of the process de- 
scribed in Example 3- IS is shown in Fig. 3-4. 
Notice that the pressure in psfa is used as the 
vertical coordinate, whereas the volume in cubic 
feet is used as the horizontal coordinate. 

In Example 3-1 5, the initial condition of the 
gas is such that the pressure is 21 17 psfa and the 
volume is 13.34 cu ft. To establish the initial 
state of the gas on the PV chart, start at the 
origin and proceed upward along the vertical 
pressure axis to the given pressure, 2117 psfa. 
Draw a dotted line parallel to the base line 
through this point and across the chart. Next, 
from the point of origin proceed to the right 
along the horizontal volume axis to the given 
volume, 13.34 cu ft. Through this point draw a 
vertical dotted line across the chart. The inter- 
section of the dotted lines at point 1 establishes 
the initial thermodynamic state of the gas. 

According to Example 3-1 5, the gas is heated 
and allowed to expand at a constant pressure 
until its volume is IS cu ft. Since the pressure 
remains the same during the process, the state 
point representing the final state of the gas must 
fall somewhere along the line of constant 
pressure already established. The exact point 
on the pressure line which represents the final 
state 2 is determined by the intersection of the 
line drawn through the point on the volume axis 
that identifies the final volume. 

In passing from the initial state 1 to the final 
state 2 the air passes through a number of 
intermediate thermodynamic states, all of which 
can be represented by points which will fall 
along line 1 to 2. Line 1 to 2, then, represents 
the path that the process will follow as the ther- 
modynamic state of the gas changes from 1 to 2, 
and is the PV diagram of the process described. 

The area of a rectangle is the product of its 



4000 





_ Final 
condition 



12 



13 ! 14 



*i 



v 2 



16 



Volume (cubic feet) 
Fig. 3-4. Pressure-volume diagram of constant 
pressure process. Crosshatched area between 
process diagram and base line represents external 
work done during the process. 



two dimensions. In Fig. 3-4, the area of the 
rectangle, 1-2-Kg-^ (crosshatched), is the 
product of its altitude P and its base (K g — Fj). 
But according to Equation 3-17, the product 
P(V t — VJ is the external work done during a 
constant pressure process. It is evident then 
that the area between the process diagram and 
the volume axis is a measure of the external 
work done during the process in foot-pounds. 
This area is frequently referred to as "the area 
under the curve." 

Figure 3-5 is a PV diagram of a constant 
volume process. Assume that the initial 
condition of the gas at the start of the process 
is such that the pressure is 2000 psfa and the 
volume is 4 cu ft. The gas is heated while its 
volume is kept constant until the pressure 
increases to 4000 psfa. The process takes place 
along the constant volume line from the initial 
condition 1 to the final condition 2. 

It has been stated that no work is done during 
a process unless the volume of the gas changes. 
Examination of the PV diagram in Fig. 3-5 will 
show that no work is indicated for the constant 
volume process. Since a line has only the 
dimension of length, there is no area between 
the process diagram and the base or volume 
axis. Hence, no work is done. 
3-25. Constant Temperature Process. 
According to Boyle's law, when a gas is com- 
pressed or expanded at a constant temperature, 
the pressure will vary inversely with the volume. 
That is, the pressure increases as the gas is 



36 PRINCIPLES OF REFRIGERATION 



4500 

4000 

3500 

3000 

32500 

£ 

| 2000 

1500 



500 




P* 


^~ Final condition 


Px 


^ — Initial condition 


— 


II 1 



3 4 5 6 7 

Volume (cubic feet) 

Fig. 3-5. Pressure-volume diagram of constant 
volume process. Since there is no area between the 
process diagram and the volume axis, there is no 
work done during a constant volume process. 



compressed and decreases as the gas is expanded. 
Since the gas will do work as it expands, if 
the temperature is to remain constant, energy 
with which to do the work must be absorbed 
from an external source (Fig. 3-26). However, 
since the temperature of the gas remains 
constant, ail of the energy absorbed by the gas 
during the process leaves the gas as work; 
none is stored in the gas as an increase in the 
internal energy. 

When a gas is compressed, work is done on 
the gas, and if the gas is not cooled during ihe 
compression, the internal energy of the gas will 
be increased by an amount equal to the work of 
compression. Therefore, if the temperature of 
the gas is to remain constant during the com- 
pression, the gas must reject to some external 
body an amount of heat equal to the amount of 
work done upon it during the compression 
(Fig. 3-2c). 

There is no change in the internal kinetic 
energy during a constant temperature process. 
Therefore, in Equation 3-12, AK is equal to zero 
and the general energy equation for a constant 
temperature process may be written 

AQ f = AW t (3-21) 

3-26. Work of an Isothermal Process. A 

PV diagram of an isothermal expansion is 



shown in Fig. 3-6. In a constant temperature 
process the pressure and volume both change in 
accordance with Boyle's law. The path followed 
by an isothermal expansion is indicated by 
line 1 to 2 and the work of the process in foot- 
pounds is represented by the area l-2-V^-V^ 
The area, 1-2-V^-V^ and therefore the work 
of the process, may be calculated by the equa- 
tion 



fF = P 1 K 1 xln-^ 



(3-22) 



where In = natural logarithm (log to the base e) 

Example 3-18. A certain weight of gas 
having an initial pressure of 2500 psfa and an 
initial volume of 2 cu ft is expanded isothermally 
to a volume of 4 cu ft. Determine: 

(a) the final pressure of the gas in psfa 

(b) the work done in heat energy units. 



Solution 
(a) By applying Boyle's 
law, Equation 3-6, 
the final pressure P a 


PiVx 

Vi 
2500 x 2 

4 
= 1250 psfa 

= 2500 x 2 
= 2500 x 2 




(b) By applying Equa- 
tion 3-22, the ex- 
ternal work of the 


x lnf 
x In 2 


process in foot- 
pounds W 


= 2500 x 2 
= 3465 ft-Ib 


x 0.693 


By Applying Equa- 
tion 3- 1 8 , the work in 
heat energy units Alf 


3465 
~ 778 
= 4.45 Btu 






Final 
condition 



Fig. 3-6. Pressure-volume diagram of constant 
temperature process. Crosshatched area represents 
the work of the process. 



THERMODYNAMIC PROCESSES 37 



Example 3-19. A certain weight of gas 
having an initial pressure of 12S0 psfa and an 
initial volume of 4 cu ft is compressed isotherm- 
ally to a volume of 2 cu ft. Determine: 

(a) the final pressure in psfa 

(b) the work done by the gas in Btu. 



Solution 




(a) By applying 
Boyle's law, 
Equation 3-6, 
the final pressure 


PiVi 

1250 x 4 

2 


^2 


= 2500 psfa 


(b) By Applying Equa- 
tion 3-22, the ex- 


= 2500 x 2 x In | 


ternal work of the 


= 2500 x 2 x In 0.5 


process in foot- 


= 2500 x 2 x -0.693 


pounds W 


= -3465 ft-lb 


By applying 
Equation 3-18, 
the work in heat 


-3465 
778 


energy units AW 


= -4.45 Btu 



Notice that the process in Example 3-19 is the 
exact reverse of that of Example 3-18. Where 
the process in Example 3-18 is an expansion, 
the process in Example 3-19 is a compression. 
Both processes occur between the same two 
conditions, except that the initial and final 
conditions are reversed. Notice also that 
whereas work is done by the gas during the 
expansion process, work is done on the gas 
during the compression process. But since the 
change of condition takes place between the 
same limits in both cases, the amount of work 
done in each case is the same. 
3-27. Heat Transferred during a Constant 
Temperature Process. Since there is no 
change in the temperature during an isothermal 
process, there is no change in the internal 
kinetic energy and AK equals zero. According 
to Equation 3-21, the heat energy transferred 
during a constant temperature process is exactly 
equal to the work done in Btu. During an 
isothermal expansion heat is transferred to the 
gas to supply the energy to do the work that is 
done by the gas, whereas during an isothermal 
compression heat is transferred from the gas so 
that the internal energy of the gas is not in- 
creased by the performance of work on the gas. 

Example 3-20. Determine the quantity of 
heat transferred to the gas during the constant 
temperature expansion described in Example 
3-18. 



Solution. From Example 
3-18, AW = 4.45 Btu 

Since, in the isothermal 
process, AW t equals AQ t , AQ t = 4.45 Btu 

Example 3-21. What is the quantity of 
heat transferred to the gas during the constant 
temperature process described in Example 3-19. 

Solution. From 
Example 3-19, AW = -4.45 Btu 

Since A W t equals AQ t , AQ t = -4.45 Btu 

Again, notice that a negative amount of heat 
is transferred to the gas, indicating that heat in 
this amount is actually given up by the gas 
during the process. 

3-28. Adiabatic Process. An adiabatic pro- 
cess is described as one wherein the gas 
changes its condition without absorbing or 
rejecting heat, as such, from or to an external 
body during the process. Furthermore, the 
pressure, volume, and temperature of the gas all 
vary during an adiabatic process, none of them 
remaining constant. 

When a gas expands adiabatically, as in any 
other expansion, the gas does external work and 
energy is required to do the work. In the 
processes previously described, the gas absorbed 
the energy to do the work from an external 
source. Since, during an adiabatic process, no 
heat is absorbed from an external source, the gas 
must do the external work at the expense of its 
own energy. An adiabatic expansion is always 
accompanied by a decrease in the temperature 
of the gas as the gas gives up its own internal 
energy to do the work (Fig. 3-7). 

When a gas is compressed adiabatically, 
work is done on the gas by an external body. 
The energy of the gas is increased in an amount 
equal to the amount of work done, and since no 
heat energy is given up by the gas to an external 
body during the compression, the heat energy 
equivalent of the work done on the gas is set up 
as an increase in the internal energy, and the 
temperature of the gas increases. 

Because no heat, as such, is transferred to or 
from the gas during an adiabatic process, AQ a is 
always zero and the energy equation for an 
adiabatic process is written as follows: 



AK a +AW a =0 



Therefore, 



(3-23) 



AW. = -AK. and AK a - -AW a 



38 PRINCIPLES OF REFRIGERATION 




Final 
condition 



Fig. 3-7. Pressure-volume diagram of adiabatic 
process. An isothermal curve is drawn in for 
comparison. 



3-29. Work of an Adiabatic Process. The 

work of an adiabatic process may be evaluated 
by the following equation: 

P 1 V 1 -P 2 V 2 



W a = 



k - 1 



(3-24) 



where k — the ratio of the specific heats of the 
gas in question, CJC V 

Example 3-22. A gas having an initial 
pressure of 2500 psfa and an initial volume of 
2 cu ft is expanded adiabatically to a volume 
of 4 cu ft. If the final pressure is 945 psfa, 
determine the external work done in heat 



energy units. 




Solution 




Cj, for air 
C„ for air 

The ratio of the specific 
heats, k 


= 0.2375 Btu/lb 
= 0.169 Btu/lb 

c, 




0.2375 




0.169 




= 1.406 


Applying Equation 3-24, 
the work of adiabatic ex- 


(2500 x 2) 
- (945 x 4) 


pansion in foot-pounds 


1406 - 1 


w a 


5000 - 3780 




0.406 




1220 




0.406 




= 3005 ft-lb 


Applying Equation 3-18, 
the work in heat energy 


3005 

~ 778 


units A W a 


= 3.86 Btu 



Example 3-23. A gas having an initial 
pressure of 945 psfa and an initial volume of 
4 cu ft is compressed adiabatically to a volume 
of 2 cu ft. If the final pressure of the air is 
2500 psfa, how much work is done in heat 
energy units? 



Solution. From 
Example 3-22, k for air 

Applying Equation 3-24, 
the work done in foot- 
pounds W a 



1.406 

(945 x 4) - 
(2500 x 2) 
1.406 - 1 

3780 - 5000 
0.406 

-1220 



Applying Equation 3-18, 
the work in heat energy 
units A(C. 



0.406 

-3005 ft-lb 
-3005 



778 
-3.86 Btu 



3-30. Comparison of the Isothermal and 
Adiabatic Processes. A comparison of the 
isothermal and adiabatic processes is of interest. 
Whenever a gas expands, work is done by the 
gas, and energy from some source is required to 
do the work. In an isothermal expansion, all 
of the energy to do the work is supplied to the 
gas as heat from an external source. Since the 
energy is supplied to the gas from an external 
source at exactly the same rate that the gas is 
doing work, the internal energy of the gas 
neither increases nor decreases and the tempera- 
ture of the gas remains constant during the 
process. On the other hand, in an adiabatic 
expansion there is no transfer of heat to the gas 
during the process and all of the work of expan- 
sion must be done at the expense of the internal 
energy of the gas. Therefore, the internal 
energy of the gas is always diminished by an 
amount equal to the amount of work done and 
the temperature of the gas decreases accordingly. 
Consider now isothermal and adiabatic 
compression processes. In any compression 
process, work is done on the gas by the com- 
pressing member, usually a piston, and an 
amount of energy equal to the amount of work 
done on the gas is transferred to the gas as work. 
During an isothermal compression process, 
energy is transferred as heat from the gas to an 
external sink at exactly the same rate that work 
is being done on the gas. Therefore, the internal 



THERMODYNAMIC PROCESSES 39 



energy of the gas neither increases nor decreases 
during the process and the temperature of the 
gas remains constant. On the other hand, 
during an adiabatic compression, there is no 
transfer of energy as heat from the gas to an 
external sink. Therefore, an amount of energy 
equal to the amount of work done on the gas is 
set up in the gas as an increase in the internal 
energy, and the temperature of the gas increases 
accordingly. 

3-34. The Polytropic Process. Perhaps the 
simplest way of denning a polytropic process is 
by comparison with the isothermal and adia- 
batic processes. The isothermal expansion, in 
which the energy to do the work of expansion is 
supplied entirely from an external source, and 
the adiabatic expansion, in which the energy to 
do the work of expansion is supplied entirely by 
the gas itself, may be thought of as the extreme 
limits between which all expansion processes will 
fall. Then, any expansion process in which the 
energy to do the work of expansion is supplied 
partly from an external source and partly from 
the gas itself will follow a path which will fall 
somewhere between those of the isothermal and 
adiabatic processes (Fig. 3-8). Such a process is 
known as a polytropic process. If during a 
polytropic' expansion most of the energy to do 
the work comes from an external source, the 
polytropic process will more nearly approach 
the isothermal. On the other hand, when the 
greater part of the energy to do the external 
work comes from the gas itself, the process 
more nearly approaches the adiabatic. 

This is true also for the compression process. 
When a gas loses heat during a compression 
process, butnotataratesumcient to maintain the 
temperature constant, the compression is poly- 
tropic. The greater the loss of heat, the closer 
the polytropic process approaches the iso- 
thermal.' The smaller the loss of heat, the closer 
the polytropic process approaches the adiabatic. 
Of course, with no heat loss, the process 
becomes adiabatic. 

The actual compression of a gas in a compres- 
sor will usually very nearly approach adiabatic 
compression. This is because the time of com- 
pression is normally very short and there is not 
sufficient time for any significant amount of heat 
to be transferred from the gas through the 
cylinder walls to the surroundings. Water 
jacketing of the cylinder will usually increase the 



Polytropic N > 1 and < K 




Fig. 3-8. Pressure-volume diagram of a polytropic 
process. Adiabatic and isothermal curves are drawn 
in for comparison. 



rate of heat rejection and move the path of the 
compression closer to the isothermal. 
3-32. PVT Relationship during Adiabatic 
Processes. Since the temperature, pressure, 
and volume all change during an adiabatic 
process, they will not vary in accordance with 
Charles' and Boyle's laws. The relationship 
between the pressure, temperature, and volume 
during an adiabatic process may be evaluated by 
the following equations : 



T t = T x x 



H' 



* -1) 



F (*-l) 






p*=Pi 



(3-25) 



(3-26) 



(3-27) 



- P '*W 


(3-28) 




(3-29) 


- v, . (3>r 


(3-30) 



Example 3-24. Air is expanded adiabati- 
cally from a volume of 2 cu ft to a volume of 



40 PRINCIPLES OF REFRIGERATION 



4 cu ft. If the initial pressure of the air is 24,000 
psfa, what is the final pressure in psfa? 



Solution. From 
Table 3-1, k for air 

Applying Equation 
3-27, the final pressure 



,1.406 



= 1.406 
= 24,000 x 

= 24,000 x 0.5 1 * 06 
= 24,000 x 0.378 
= 9072 psfa 



Example 3-25. Air is expanded adia- 
batically from a volume of 2 cu ft to a volume 
of 4 cu ft. If the initial temperature of the air 
is 600° R, what is the final temperature in 
degrees Rankine? 



Solution. From 
Table 3-1, k for air 

Applying Equation 
3-25, the final tempera- 
ture T 2 



= 1.406 
= 600x 



(2)« 



.406-1) 



= 600x 



= 600x 



(4)(1.406- 

(2)0.406 

(4)0.408 

1.325 



1.756 
= 600 x 0.755 
= 453" R 

Example 3-26. Air is expanded adia- 
batically from an initial pressure of 24,000 psfa 
to a final pressure of 9072 psfa. If the initial 
temperature is 600° R, what is the final tem- 
perature? 



Solution. From 
Table 3-1, k for 
air 

Applying 
Equation 3-26, 
the final tempera- 
ture T 9 



= 1.406 



= 600 x 



1.406-1 

9072 \ i«o« 



/ 9072 \ ** 
\24,000/ 



124 000/ 
= 600 x (0.378)< 0M »» 
= 600 x 0.755 
= 453° R 



3-33. Exponent of Polytropic Expansion 
and Compression. The presssure-tempera- 
ture-volume relationships for the polytropic 
process can be evaluated by Equations 3-25 
through 3-30, except that the polytropic ex- 
pansion or compression exponent n is substi- 
tuted for k. Too, the work of a polytropic 
process can be determined by Equation 3-24 if 
n is substituted for k. 

The exponent n will always have a value 
somewhere between 1 and k for the particular 



gas ' undergoing the process.* Usually, the 
value of « must be determined by actual test of 
the machine in which the expansion or com- 
pression occurs. In some instances average 
values of n for some of the common gases 
undergoing changes under more or less standard 
conditions are given in tables. If the values of 
two properties are known for both initial and 
final conditions, the value of n may be calcu- 
lated. The following sample equation shows 
the relationship: 

log (Pj/Pj,) 

-^ (3-31) 



« = 



logC^/^) 



Example 3-27." Air, having an initial 
pressure of 24,000 psfa and an initial tempera- 
ture of 600° R, is expanded polytropically from 
a volume of 2 cu ft to a volume of 4 cu ft. If 
the exponent of polytropic expansion is 1.2, 
determine: 

(a) The weight of the air in pounds 

(6) The final pressure in psfa 

(c) The final temperature in degrees Rankine 

(d) The work done by the gas in Btu 

(e) The increase in internal energy 
(/) The heat transferred to the gas. 



Solution 
(a) From Table 3-1, 

R for air 

Rearranging and 
applying Equation 
3-10, the weight of 
air M 



= 53.3 

RT 
_ 24,000 x 2 

53.3 x 600 
= 1.5 lb 



* The value of n depends upon the specific heat 
of the gas during the process. Since the specific 
heat may take any value, it follows that theoretically 
n may have any value. In actual machines, however, 
n will nearly always have some value between 1 
and A:. 

Broadly- defined, a polytropic process is any 
process during which the specific heat remains 
constant. By this definition, all five processes 
discussed in this chapter are polytropic processes. 
It is general practice today to restrict the term 
polytropic to mean only those processes which follow 
a path falling between those of the isothermal and 
adiabatic processes. The exponents of isothermal 
and adiabatic expansions or compressions are 1 
and k, respectively. Hence, the value of n for the 
polytropic process must fall between 1 and k. The 
closer the polytropic process approaches the 
adiabatic, the closer n will approach k. 



(b) Applying Equation 
3-27, the final 


- 24,000 xg)" 


pressure P a 


= 24,000 x (0.SY* 




- 24,000 x 0.435 




= 10,440 psfa 


(c) Applying Equation 


(2)U.»-i) 

— 600 x 


3-25, the tempera- 


""" * (4)<i.«-i) 


ture T t 


-«0x3K 




(4) o.» 




,„ 1.149 




= 600 x — — 




1.32 




= 522°R 


(d) Applying Equation 


(24,000 x 2) 


3-24, the work 


- (4 x 10,440) 


done W 


1.2 - 1 




_ 48,000 - 41,760 




0.2 




6240 








0.2 




= 31,200 ft-lb 


Applying Equation 


31,200 


3-18, the work in 


778 


BtuAW 


= 40.10 Btu 


(e) From Table 3-1, 


= 0.169 Btu/lb 


C„ for air 




Applying Equation 


= 1.5 x 0.169 x 


3-15, the increase 


(522 - 600) 


in internal energy 


=■ 1.5 x 0.169 x 


AK 


(-78) 




= -19.77 Btu 


(/) Applying Equation 


= AJSr + AfF 


3-9, the heat 




energy trans- 


= -19.77 + 40.10 


ferred to the gas AQ 


= 20.33 Btu 



Notice in Example 3-27 that the work done 
by the air in the polytropic expansion is equiva- 
lent to 40.10 Btu. Of this amount, 20.33 Btu 
is supplied from an external source, whereas the 
other portion, 19.77 Btu, is supplied by the gas 
itself, thereby reducing the internal kinetic 
energy by this amount. 

PROBLEMS 

1. Three pounds of air occupy a volume of 
24 cu ft. Determine: 

(a) The density of the air. Arts. 0.125 lb/cu ft 

(b) The specific volume. Am. 8 cu ft/lb 



THERMODYNAMIC PROCESSES 41 

2. The volume of a certain weight of air is kept 
constant while the temperature of the air is 
increased from 55° F to 100° F. If the initial 
pressure is 25 psig, what is the final pressure of 
the air in psig ? Am. 28 .47 psig 

3. A certain weight of air confined in a con- 
tainer is cooled from 150° F to 70° F. If the 
initial pressure of the air is 36.3 psig, what is the 
final pressure of the air in psig ? Am. 29.6 psig 

4. One pound of air at atmospheric pressure 
has a volume of 13.34 cu ft at a temperature 
of 70° F. If the air is passed across a heat 
exchanger and is heated to a temperature of 
150° F while its pressure is kept constant, what 
is the final volume of the air ? Ans. 15.35 cu ft 

5. A cylinder of oxygen has a volume of 5 cu ft. 
A gage on the cylinder reads 2200 psi. If the 
temperature of the oxygen is 85° F, what is the 
weight of the oxygen in the cylinder? 

Am. 60.6 lb 

6. In Problem 4, determine: 

(a) The work done by the air during the 
heating. Am. 4254.8 ft-lb or 5.47 Btu 

(b) The increase in the internal kinetic energy. 

Am. 13.52 Btu 

(c) The quantity of heat transferred to the air. 

Am. 19 Btu 

7. A certain weight of air having an initial 
volume of 0.1334 cu ft and an initial tempera- 
ture of 70° F is drawn into the suction side of 
an air compressor. If the air enters the cylinder 
at standard atmospheric pressure and is com- 
pressed isothermally to a final pressure of 150 
psia, determine: 

(a) The weight of air in the cylinder at the 
start of the compression stroke. 

Am. 0.01 lb 

(b) The final temperature of the air in degrees 
Rankine. Am. 530° R 

(c) The volume of the air at the end of the 
compression stroke. Am. 0.0131 cu ft 

(d) The work of compression in Btu. 

Am. 0.843 Btu 

(e) The increase or decrease in internal energy. 

Am. None 

(/) The energy transferred to the gas during 

the compression. Am. —0.843 Btu 

8. Assume that the air in Problem 7 is com- 
pressed adiabatically rather than isothermally. 
Compute: 

(a) The final temperature of the air in degrees 
Rankine. Am. 1038° R 

(b) The volume of the air at the end of the 
compression stroke. Am. 0.0256 cu ft 

(c) The work of compression in Btu. 

Am. 0.86 Btu 



42 PRINCIPLES OF REFRIGERATION 

(d) The increase in the internal kinetic energy. 

Arts. 0.86 Btu 
(c) The heat energy transferred to or from 
the gas during the compression. 

Ans. None 

9. Assuming that the air in Problem 7 is com- 
pressed polytropically rather than isothermally. 
If n equals 1.2, compute: 
(«) The final temperature of the air in degrees 
Rankine. Ans. 119 A" R 



(b) The volume of the air at the end of the 
compression stroke. Ans. 0.0192 cu ft 

(c) The work of compression is Btu. 

Ans. 0.8S Btu 

(d) The decrease in the internal kinetic energy. 

Ans. 0.45 Btu 

(e) The heat energy transferred from the gas 
during the compression. Ans. 0.40 Btu 

10. Compare the results of Problems 7, 8, 
and 9. 



4 

Saturated and 

Superheated 

Vapors 



4-1. Saturation Temperature. When the 
temperature of a liquid is raised to a point such 
that any additional heat added to the liquid will 
cause a part of the liquid to vaporize, the 
liquid is said to be saturated. Such a liquid is 
known as a saturated liquid and the tempera- 
ture of the liquid at that condition is called the 
saturation temperature (Sections 2-31 and 2-32). 
4-2. Saturated Vapor. The vapor ensuing 
from a vaporizing liquid is called a saturated 
vapor as long as the temperature and pressure 
of the vapor are the same as those of the satur- 
ated liquid from which it came. A saturated 
vapor may be described also as a vapor at a 
temperature such that any further cooling of 
the vapor will cause a portion of the vapor to 
condense and thereby resume the molecular 
structure of the liquid state. It is important 
to understand that the saturation temperature 
of the liquid (the temperature at which the liquid 
will vaporize if heat is applied) and the satura- 
tion temperature of the vapor (the temperature 
at which the vapor will condense if heat is 
removed) are the same for any given pressure 
and that the liquid cannot exist as a liquid at any 
temperature above its saturation temperature, 
whereas a vapor cannot exist as a vapor at any 
temperature below its saturation temperature.* 

* Under certain conditions it is possible to "super- 
cool" water vapor momentarily below its saturation 
temperature. However, this is a very unstable 
condition and cannot be maintained except momen- 
tarily. 



For example, in Fig. 4-1 , the water in the heated 
vessel is saturated and is vaporizing at 212° F as 
the latent heat of vaporization is supplied by the 
burner. The water vapor (steam) rising from 
the water is saturated and remains at the satura- 
tion temperature (212° F) until it reaches the 
condenser. As the saturated vapOr gives up 
heat to the cooler water in the condenser, it 
condenses back into the liquid state. Since 
condensation occurs at a constant temperature, 
the water resulting from the condensing vapor 
is also at 212° F. The latent heat of vapor- 
ization, absorbed as the water vaporizes into 
steam, is given up by the steam as the steam 
condenses back into water. 
4-3. Superheated Vapor. A vapor at any 
temperature above its saturation temperature 
is a superheated vapor (Section 2-34). If, after 
vaporization, a vapor is heated so that its 
temperature is raised above that of the vapor- 
izing liquid, the vapor is said to be superheated. 
In order to superheat a vapor it is necessary to 
separate the vapor from the vaporizing liquid as 
shown in Fig. 4-2. As long as the vapor remains 
in contact with the liquid it will be saturated. 
This is because any heat added to a liquid- 
vapor mixture will merely vaporize more liquid 
and no superheating will occur. 

Before a superheated vapor can be condensed, 
the vapor must be de-superheated, that is, the 
vapor must first be cooled to its saturation 
temperature. Heat removed from a super- 
heated vapor will cause the temperature of the 
vapor to decrease until the saturation tempera- 
ture is reached. At this point, any further 
removal of heat will cause a part of the vapor to 
condense. 

4-4. Subcooled Liquid. If, after condensa- 
tion, a liquid is cooled so that its temperature 
•is reduced below the saturation temperature, the 
liquid is said to be subcooled. Thus, a liquid at 
any temperature below the saturation tempera- 
ture and above the fusion point is a subcooled 
liquid. 

4-5. The Effect of Pressure on the Satura- 
tion Temperature. The saturation tempera- 
ture of a liquid or a vapor varies with the 
pressure. Increasing the pressure raises the 
saturation temperature and decreasing the pres- 
sure lowers the saturation temperature. For 
example, the saturation temperature of water at 
atmospheric pressure (0 psig or 14.7 psia) is 



43 



44 PRINCIPLES OF REFRIGERATION 



212° F. If the pressure over the water is in- 
creased from psig to 5.3 psig (20 psiaX the 
saturation temperature of the water increases 
from 212° F to 228° F. On the Other hand, if the 
pressure over the water is reduced from 14.7 
psia to 10 psia, the new saturation temperature 
of the water will be 193.2° F. Figure 4-3 is a 



the water at atmospheric pressure is 212° F, 
the temperature of the water will rise as the 
water is heated until it reaches 212° F. At this 
point, if the heating is continued, the water will 
begin to vaporize. Soon the space above the 
water will be filled with billions and billions of 
water vapor molecules darting about at high 



Condenser 
water out""^ 



Saturated steam 
at212"F -* 



Heat added 



Steam gives up heat to 

, cold water in condenser 

/ and condenses into water 



.{ ^Con densed steam leaving 
condenser-212°F 




Cold water in 



Fig. 4VI. Saturated vapor. 



graphical representation of the relationship 
between the pressure and the saturation tempera- 
ture of water. 

To illustrate the effect of pressure on the 
saturation temperature of a liquid, assume that 
water is confined in a closed vessel which is 
equipped with a throttling valve at the top 
(Fig. 4-4a). A compound gage is used to deter- 
mine the pressure exerted in the vessel and two 
thermometers are installed so that one records 
the temperature of the water and the other the 
temperature of the vapor over the water. With 
the throttling valve wide open, the pressure 
exerted over the water is atmospheric (0 psig or 
14.7 psia). Since the saturation temperature of 



velocities. Some of the vapor molecules will fall 
back into the water to become liquid molecules 
again, whereas others will escape through the 
opening to the outside and be carried away by 
air currents. If the opening at the top of the 
vessel is of sufficient size to allow the vapor to 
escape freely, the vapor will leave the vessel at 
the same rate that the liquid is vaporizing. That 
is, the number of molecules which are leaving 
the liquid to become vapor molecules will be 
exactly equal to the number of vapor molecules 
which are leaving the space, either by escaping 
to the outside or by falling back into the liquid. 
Thus, the number of vapor molecules and the 
density of the vapor above the liquid will 



SATURATED AND SUPERHEATED VAPORS 45 



Fig. 4-2. Superheated vapor. 




« r- : iOiA-.t!j>.^. -^j^~ -± Jt.i-^x4j— ^ ^.j ^Ls ^Steam superheated 

in superheater 

Saturated steam 



tint 

Heat added 

remain constant and the pressure exerted by the 
vapor will be equal to that of the atmosphere 
outside of the vessel. 

Under this condition the water vapor ensuing 
from the vaporizing liquid will be saturated, 
that is, its temperature and pressure will be the 
same as that of the water, 212° F and 14.7 psia. 
The density of the water vapor at that tempera- 
ture and pressure will be 0.0373 lb/cu ft and its 
specific volume will be 1 /0.0373 or 26.8 ft*/lb. 

Regardless of the rate at which the liquid is 
vaporizing, as long as the vapor is allowed to 
escape freely to the outside so that the pressure 
and density of the vapor over the liquid does 
not change, the liquid will continue to vaporize 
at 212° F. 




/T7-X 



TfT"/ 5 " 

H eat Superheated 



added 



steam 



Suppose that the throttling valve is partially 
closed so that the escape of the vapor from the 
vessel is impeded somewhat. For a time the 
equilibrium will be disturbed in that the vapor 
will not be leaving the vessel at the same rate the 
liquid is vaporizing. The number of vapor 
molecules in the space above the liquid will 
increase, thereby increasing the density and the 
pressure of the vapor over the liquid and raising 
the saturation temperature. 

If it is assumed that the pressure of the vapor 
increases to 5.3 psig (20 psia) before equilibrium 
is again established, that is, before the rate at 
which the vapor is escaping to the outside is 
exactly equal to the rate at which the liquid is 
vaporizing, the saturation temperature will be 



DUU 

500 
-.400 

re 
'3» 
o. 

Is 300 
a. 200 

100 
14.7 



























































































































Fig. 4-3. Variation in the 
saturation temperature of 
water with changes in pressure. 



50 100 150 200| 250 300 350 400 450 500 
212 
Saturation temperature (*F) 



46 PRINCIPLES OF REFRIGERATION 









xtt v' Valve wit,e °P en 






« /S^ Steam <„'•> 
Density-0.0373 Ib/cu ft 




: Z z _Water_ H>: : >^^ 




Valve partially closed 

20 psia 



►212'F 



^ 



^i'-'i- Steam 
Density-0.0498 Ib/cu ft 




>228'F 



(a) 



<b) 



Fig. 4-4. 



228° F, the density of the vapor will be 0.0498 
lb/cu ft, and 1 lb of vapor will occupy a volume 
of 20.08 cu ft. This condition is illustrated in 
Fig. 4-46. 

By comparing the condition of the vapor in 
Fig. 4-4A with that of the vapor in Fig. 4-4a, it 
will be noted that the density of the vapor is 
greater at the higher pressure and saturation 
temperature. Furthermore, it is evident that the 
pressure and the saturation temperature of 
liquid or vapor can be controlled by regulating 
the rate at which the vapor escapes from over 
the liquid. 

In Fig. 4-4«, the rate of vaporization will have 
little or no effect on the pressure and saturation 
temperature because the vapor is allowed to 
escape freely so that the density and pressure of 
the vapor over the liquid will neither increase 
nor decrease as the rate of vaporization is 
changed. On the other hand, in Fig. 4-46, any 
increase in the rate of vaporization will cause an 
increase in the density and pressure of the vapor 
and result in an increase in the saturation 
temperature. The reason is that any increase 
in the rate of vaporization will necessitate the 
escape of a greater quantity of vapor in a given 
length of time. Since the size of the vapor outlet 
is fixed by the throttling action of the valve, the 
pressure of the vapor in the vessel will increase 
until the pressure difference between the inside 



and outside of the vessel is sufficient to allow 
the vapor to escape at a rate equal to that at 
which the liquid is Vaporizing. The increase in 
pressure, of course, results in an increase in the 
saturation temperature and in the density of the 
vapor. Likewise, any decrease in the rate of 
vaporization will have the opposite effect. The 
pressure and density of the vapor over the liquid 
will decrease and the saturation temperature 
will be lower. 

Assume now that the throttling valve on the 
container is again opened completely, as in 
Fig. 4-4a, so that the vapor is allowed to escape 
freely and unimpeded from over the liquid. The 
density and pressure of the vapor will decrease 
until the pressure of the vapor is again equal to 
that of the atmosphere outside of the container. 
Since the saturation temperature of water at 
atmospheric pressure is 212° F and since a liquid 
cannot exist as a liquid at any temperature above 
its saturation temperature corresponding to its 
pressure, it is evident that the water must cool 
itself from 228° F to 212° F at the instant that 
the pressure drops from 20 psia to atmospheric 
pressure. To accomplish this cooling, a portion 
of the liquid will "flash" into a vapor. The 
latent heat necessary to vaporize the portion of 
the liquid that flashes into the vapor state is 
supplied by the mass of the liquid and, as a 
result of supplying the vaporizing heat, the 



SATURATED AND SUPERHEATED VAPORS 47 



temperature of the mass of the liquid will be 
reduced to the new saturation temperature. 
Enough of the liquid will vaporize to provide 
the required amount of cooling. 
4-6. Vaporization. The vaporization of a 
liquid may occur in two ways: (1) by evapora- 
tion and (2) by ebullition or "boiling." The 
vaporization of a liquid by evaporation occurs 
only at the free surface of the liquid and may 
take place at any temperature below the satura- 
tion temperature. On the other hand, ebullition 
or boiling takes place both at the free surface 
and within the body of the liquid and can occur 
only at the saturation temperature. Up to this 
point, only ebullition or boiling has been 
considered. 

4-7. Evaporation. Evaporation is taking 
place continually and the fact that water 
evaporates from lakes, rivers, ponds, clothes, 
etc., is sufficient evidence that evaporation can 
and does occur at temperatures below the 
saturation temperature. Any liquid open to the 
atmosphere, regardless of its temperature, will 
gradually evaporate and be diffused into the 
air. 

The vaporization of liquids at temperatures 
below their saturation temperature can be 
explained in this manner. The molecules of a 
liquid are in constant and rapid motion, their 
velocities being determined by the temperature 
of the liquid. In the course of their movements 
the molecules are continually colliding with one 
another and, as a result of these impacts, some 
of the molecules of the liquid momentarily 
attain velocities much higher than the average 
velocity of the other molecules of the mass. 
Thus, their energy is much greater than the 
average energy of the mass. If this occurs 
within the body of the liquid, the high velocity 
molecules quickly lose their extra energy in 
subsequent collisions with other molecules. 
However, if the molecules attaining the higher 
than normal velocities are near the surface, they 
may project themselves from the surface of the 
liquid and escape into the air to become vapor 
molecules. (Fig. 4-5). The molecules so 
escaping from the liquid are diffused throughout 
the air. They occupy the relatively large spaces 
which exist between the molecules of the air and 
become a part of the atmospheric air. 
4-8. Rate of Vaporization. For any given 
temperature, some liquids will evaporate faster 



High energy molecules 
escape from surface 
S" of liquid and become 
f vapor molecules 




Water 



Fig. 4-5. Evaporation from surface of a liquid. 



than others. Liquids havingthelowest "boiling" 
points, that is, the lowest saturation temperature 
for a given pressure, evaporate at the highest 
rate. However, for any particular liquid, the 
rate of vaporization varies with a number of 
factors. In general, the rate of vaporization 
increases as the temperature of the liquid 
increases and as the pressure over the liquid 
decreases. Evaporation increases also with the 
amount of exposed surface. Furthermore, it 
will be shown later that the rate of evaporation 
is dependent on the degree of saturation of the 
vapor which is always adjacent to and above 
the liquid. 

4-9. The Cooling Effect of Evaporation. 
Since it is the higher velocity molecules (those 
having the most energy) which escape from the 
surface of an evaporating liquid, it follows that 
the average energy of the mass is thereby 
reduced and the temperature of the mass 
lowered. Whenever any portion of a liquid 
vaporizes, an amount of heat equal to the latent 
heat of vaporization must be absorbed by that 
portion, either from the mass of the liquid, 
from the surrounding air, or from adjacent 
objects. Thus, the energy and temperature of 
the mass are reduced as it supplies the latent 
heat of vaporization to that portion of the 
liquid which vaporizes. The temperature of the 
mass is reduced to a point slightly below that of 
the surrounding media and the temperature 
difference so established causes heat to flow 
from the surrounding media into the mass of the 
liquid. The energy lost by the mass during 
vaporization is thereby replenished and evapora- 
tion becomes a continuous process as long as 



48 PRINCIPLES OF REFRIGERATION 



Molecules escaping from 
surface are carried away 
by air so that evaporation 
is continuous 




Water temperature 
slightly below air >. 
temperature > 



Vapor pressure 
0.74 in. Hg 
Absolute density 
0.001148 Ib/cu ft 




Molecules cannot 
escape-fall back 
into liquid to replace 
those leaving 



Saturated vapor 

70° F 
Saturated liquid 



(a) 



(b) 



Vapor pressure- 
1.13 in. Hg 
Absolute density- 
0.001570 Ib/cu ft 



Saturated vapor 




Vapor pressure— 
0.52 in. Hg 
Absolute density— 
60' F / 0.000828 

Saturated vapor 



60° F 

Saturated liquid 




(c) 



(d) 



Fig. 4-6. 



any of the liquid remains. The vapor resulting 
from evaporation is diffused into and carried 
away by the air. 

4-10. Confined Liquid-Vapor Mixtures. 
When a vapor is confined in a container with a 
portion of its own liquid, both the vapor and the 
liquid will be saturated. To illustrate, assume 
that an open container is partially filled with 
water and is stored where the ambient tempera- 
ture is 70° F (Fig. 4-6a). The water will be 
evaporating at 70° F and, as described in the 
previous section, the vapor molecules leaving 
the liquid will be diffused into the surrounding 
air so that evaporation will continue until all 
of the liquid is evaporated. However, if a 
tightly fitting cover is placed over the container, 
the vapor molecules will be unable to escape to 
the outside and they will collect above theliquid. 
Soon the space above the liquid will be so filled 
with vapor molecules that there will be as many 
molecules falling back into the liquid as there 
are leaving the liquid. A condition of equi- 
librium will be attained, the vapor will be 



saturated, and no further evaporation will 
occur. The energy of the liquid will be increased 
by the vapor molecules which are returning to 
the liquid in exactly the same amount that it is 
diminished by the molecules that are leaving. 
Since no further cooling will take place by 
evaporation, the liquid will assume the tempera- 
ture of the surrounding air and' heat transfer 
will cease. (Fig. 4-66). 

If, at this point, the ambient temperature 
rises to, say, 80° F, heat transfer will again take 
place between the surrounding air and the 
liquid. The temperature and average molecular 
velocity of the liquid will be increased and 
evaporation will be resumed. The number of 
molecules leaving the liquid will again be 
greater than the number returning and the 
density and pressure of the vapor above the 
liquid will be increased. As the density and 
pressure of the vapor increase, the saturation 
temperature of the liquid increases. Eventually, 
when the saturation temperature reaches 80° F 
and is equal to the ambient temperature, no 



further heat transfer will occur and evaporation 
will cease. Equilibrium will have again been 
established. The density and pressure of the 
vapor will be greater than before, the saturation 
temperature of the liquid-vapor mixture will be 
higher, and there will be more vapor and less 
liquid in the container than previously (Fig. 
4-6c). 

Suppose now that the ambient temperature 
falls to 60° F. When this occurs, heat will flow 
from the 80° F liquid-vapor mixture to the 
cooler surrounding air. As the liquid-vapor 
mixture loses heat to the surrounding air, its 
temperature and average molecular velocity will 
be decreased and many of the vapor molecules, 
lacking sufficient energy to remain in the vapor 
state, will fall back into the liquid and resume 
the molecular arrangement of the liquid state; 
that is, a part of the vapor will condense. The 
density and pressure of the vapor will be 
diminished and the saturation temperature of 
the mixture will be reduced. When the satura- 
tion temperature of the mixture falls to 60° F 
it will be the same as the ambient temperature 
and no further heat flow will occur. Equilibrium 
will have been established and the number of 
molecules re-entering the liquid will exactly 
equal those which are leaving. At this new 
condition, the density and pressure of the vapor 
will be less than before, the saturation tempera- 
ture will be lower, and since a part of the vapor 
condensed into liquid, there will be more liquid 
and less vapor comprising the mixture than at 
the previous condition (Fig. 4-6d). 
4-11. Sublimation. It is possible for a 
substance to go directly from the solid state 
to the vapor state without apparently passing 
through the liquid state. Any solid substance 
will sublime at any temperature below its 
fusion temperature. Sublimation takes place 
in a manner similar to evaporation, although 
much slower, in that the higher velocity 
molecules near the surface escape from the mass 
into the surrounding air and become vapor 
molecules. One of the most familiar examples 
of sublimation is that of solid C0 2 (dry ice), 
which, at normal temperatures and pressures, 
sublimes directly from the solid to the vapor 
state. Damp wash frozen on the line in the 
winter time will sublime dry. During freezing 
weather ice and snow will sublime from streets 
and sidewalks, etc. 



SATURATED AND SUPERHEATED VAPORS 49 

4-12. Condensation. Condensation of a 
vapor may be accomplished in several ways: (1) 
by extracting heat from a saturated vapor, (2) 
by compressing the vapor while its temperature 
remains constant, or (3) by some combination 
of these two methods. 

4-13. Condensing by Extracting Heat from 
a Saturated Vapor. A saturated vapor has 
been previously described as one at a condition 
such that any further cooling will cause a part 
of the vapor to condense. This is because a 
vapor cannot exist as a vapor at any tempera- 
ture below its saturation temperature. When 
the vapor is cooled, the vapor molecules cannot 
maintain sufficient energy and velocity to over- 
come the attractive forces of one another and 
remain as vapor molecules. Some of the 
molecules, overcome by the attractive forces, 
will revert to the molecular structure of the 
liquid state. When condensation occurs while 
the vapor is confined so that the volume remains 
constant, the density and pressure of the vapor 
will decrease so that there is a decrease in the 
saturation temperature. If, as in a vapor 
condenser (Fig. 4-1), more vapor is entering 
the vessel as the vapor condenses and drains 
from the vessel as a liquid, the density, pressure, 
and saturation temperature of the vapor will 
remain constant and condensation will continue 
as long as heat is continuously extracted from 
the vapor. 

4-14. Condensing by Increasing the Pres- 
sure at a Constant Temperature. When a 
vapor is compressed at a constant temperature, 
its volume diminishes and the density of the 
vapor increases as the molecules of the vapor 
are forced into a smaller volume. The satura- 
tion temperature of the vapor increases as the 
pressure increases until a point is reached where 
the saturation temperature of the vapor is equal 
to the actual temperature of the vapor. When 
this occurs, the density of the vapor will be at a 
maximum value for that condition, and any 
further compression will cause a part of the 
vapor to assume the more restrained molecular 
structure of the liquid state. Thereafter, 
condensation will continue as long as com- 
pression continues so that the density and 
pressure of the remaining vapor cannot be 
further increased. If the temperature of the 
vapor is to remain constant, heat must be 
removed from the vapor during the compression 



50 PRINCIPLES OF REFRIGERATION 



(Section 3-25). If heat is not removed from the 
vapor, the temperature of the vapor will 
increase and condensation will not occur. 

A careful analysis of Sections 4-13 and 4-14 
will show that in either case the vapor is 
brought to a saturated condition before conden- 
sation begins and that heat is removed from 
the vapor in order to bring about condensation. 
Furthermore, the vapor is saturated in each 
case only when the saturation temperature and 
the actual temperature of the vapor are the 
same. 

In Section 4-13, heat is removed from the 
vapor at a constant pressure until the tempera- 
ture of the vapor falls to the saturation tempera- 
ture corresponding to its pressure, whereupon 
the continued removal of heat causes a part of 
the vapor to condense. 

In Section 4-14, the pressure of the vapor 
is increased while the temperature of the vapor 
remains constant until the saturation tempera- 
ture of the vapor corresponding to the increased 
pressure is equal to the actual temperature of 
the vapor. In both cases, since the vapor must 
give up the latent heat of vaporization in order 
to condense, heat must be removed from the 
vapor. 

4-15. Critical Temperature. The tempera- 
ture of a gas may be raised to a point such 
that it cannot become saturated regardless of 
the amount of pressure applied. The critical 
temperature of any gas is the highest tempera- 
ture the gas can have and still be condensible 
, by the application of pressure. The critical 
temperature is different for every gas. Some 
gases have high critical temperatures while the 
^critical" temperatures of others are relatively 
low. For example, the critical temperature of 
water vapor is 706° F, whereas the critical 
temperature of air is approximately -225° F. 
4-16. Critical Pressure. Critical pressure is 
the lowest pressure at which a substance can 
exist in the liquid state at its critical tempera- 
ture; ihit is, it is the saturation pressure at the 
critical temperature. 

44J. Igriportant Properties of Gases and 
Vapors. Although a gas or vapor has many 
properties, only six are of particular importance 
in the study of refrigeration. These are pressure, 
temperature, volume, enthalpy, internal energy, 
and entropy. Pressure, temperature, and 
volume are called measurable properties be- 



cause they can actually be measured. Enthalpy, 
internal energy, and entropy cannot be measured. 
They must be calculated and are therefore 
known as calculated properties. 

Pressure, temperature, volume, and internal 
energy have already been discussed to some 
extent. A discussion of enthalpy and entropy 
follows. 

4-18. Enthalpy. Enthalpy is a calculated 
property of matter which is sometimes loosely 
defined as "total heat content." More specifi- 
cally, the enthalpy H of a given mass of material 
at any specified thermodynamic condition is an 
expression of the total heat which must be 
transferred to the material to bring the material 
to the specified condition from some initial 
condition arbitrarily taken as the zero point of 
enthalpy. 

Whereas the total enthalpy H represents the 
enthalpy of M pounds, the specific enthalpy 
h is the enthalpy of 1 lb. Since it is usually the 
specific enthalpy rather than the total enthalpy 
which is of interest, hereafter in this text the 
term enthalpy shall be used to mean specific 
enthalpy, h, the enthalpy of 1 lb. 

Since little is known about the specific heat or 
the other properties of materials at low tempera- 
tures, it is not possible to determine absolute 
values for the calculated properties. For this 
reason, values for the calculated properties 
must be determined from some arbitrarily 
selected zero point rather than from absolute 
zero.* For example, the zero point of enthalpy 
for water and its vapor, steam, is taken as 
water at 32 a F under atmospheric pressure. 
The enthalpy of 1 lb of water at 60° F then is the 
total amount of heat which must be transferred 
to the water in order to raise the temperature of 
the water from 32° F to 60° F. According to 
Equation 2-9, this is 28 Btu (1 x 1 x 28). 
Hence, based on the assumption that the 
enthalpy of water is zero at 32° F, the enthalpy 
of water at 60° F is 28 Btu/lb. 

Mathematically, enthalpy is defined as 



Pv 



(4-1) 



* Since it is required to know the change in the 
enthalpy of the working fluid during a process, 
rather than the absolute enthalpy at some par- 
ticular condition, the fact that absolute enthalpy 
cannot be calculated is of little consequence. 



SATURATED AND SUPERHEATED VAPORS 51 



Standard atmospheric pressure 
2105 psfa (14.6% x 144) 



Fig. 4-7. Pressure-volume 
diagram showing the external 
work done by fluid expansion 
as I lb of water is vaporized 
at atmospheric pressur 
approximately 59,000 ft-lb. 



« P 



S 

a. 



i s- Final condition 




Specific volume of 
1 lb of steam at 212° 
26.8 cu ft/lb 



Volume cubic feet 



where h = the specific enthalpy in Btu/lb 

u = the specific internal energy in Btu/lb 
P = the pressure in psfa 
v = the specific volume in cubic feet 
J = the mechanical energy equivalent 
It has been demonstrated (Section 3-12) that 
all the heat transferred to a fluid is not neces- 
sarily stored in the fluid as an increase in the 
internal energy of the fluid. In many cases, 
some part or all or the heat transferred to the 
fluid passes through the fluid and leaves the 
fluid as work. In Equation 4-1, that part of the 
transferred energy which is stored in the fluid 
as an increase in the internal energy is repre- 
sented by the term u, whereas that part of the 
transferred heat which leaves the fluid as work 
is represented by the term PvjJ. Notice that, 
although the energy represented by the term 
PvjJ, does not increase the internal energy of the 
fluid and is not stored in the fluid, it neverthe- 
less represents energy which must be trans- 
ferred to the fluid in order to bring the fluid 
to the specified condition from the initial 
condition at the arbitrarily selected zero point of 
enthalpy. Furthermore, even though the 
external work energy is not stored in the fluid, 
it must pass back through the fluid and be given 
up by the fluid as the fluid returns to the initial 
condition. 

Consider, for example, the vaporization of 1 
lb of water into steam at 212° F under atmos- 
pheric pressure. The volume of 1 lb of water at 
212° F is 0.01670 cu ft whereas the volume of 1 
lb of steam at 212° F is 26.82 cu ft. Hence, the 
fluid expands from a volume of 0.0167 cu ft to a 
volume of 26.82 cu ft during the vaporization 
thereby doing work in expanding against the 
pressure of the atmosphere. 



The enthalpy of vaporization (latent heat of 
vaporization) of water at 212° F is 970.4 Btu. Of 
this amount, only 897.6 Btu actually increases 
the internal energy and represents energy in 
storage in the vapor. The other 72.8 Btu leaves 
the vapor as the work of expansion and is 
represented by the term PvjJ. A PV diagram of 
the vaporization process is shown in Fig. 4-7. 
4-19. Entropy. Entropy, like enthalpy, is a 
calculated property of matter. The entropy S 
of a given mass of material at any specified 
condition is an expression of the total heat 
transferred to the material per degree of abso- 
lute temperature to bring the material to that 
condition from some initial condition taken as 
the zero of entropy. 

Since it is not possible to calculate the 
absolute value of entropy, entropy values, like 
those of enthalpy, are based on an arbitrarily 
selected zero point. The zero points of entropy 
and enthalpy are the same for any one fluid. 
Hence, for water and its vapor, steam, the zero 
point of entropy is taken as water at 32° F. 

Again, as in the case of enthalpy, it is the 
specific entropy s rather than the total entropy 
S which is useful. Therefore, in this book, the 
term entropy shall be used to mean specific 
entropy s rather than the total entropy S. 

It has been shown (Section 3-22) that the 
mechanical energy or work of a process can be 
expressed as the product of the change in 
volume and the average absolute pressure. 
Likewise, it is often convenient to express the 
heat energy transferred during a process as the 
product of two factors. The concept of entropy 
makes this possible. The heat energy trans- 
ferred during a process can be expressed as the 
product of the change in entropy and the 



52 PRINCIPLES OF REFRIGERATION 



■672*R(212'F + 460) 




Fig. 44 



Specific entropy of 
saturated steam at 212° F 



g]/ 1.7566 Btu/lb/°R 



Entropy (Btu/lb/'R) 



average absolute temperature. * Mathematically 
the relationship is expressed by the following 
equations: 

AQ = As x T m (4-2) 



a A <2 

* m 

AC 

T = — — 
lm As 



(4-3) 
(4-4) 



where AQ = the heat energy transferred in Btu 
As = the change in entropy in Btu per 

pound per ° R 
T m = the average absolute temperature 
in°R 

On a pressure-volume diagram (Fig. 4-7), the 
"area under the curve," which is the product 
of the change in volume and the average abso- 
lute pressure, represents the work of the process. 
Similarly, on a temperature-entropy diagram 
(temperature plotted against entropy), the 
"area under the curve," which is the product 
of the entropy change and the average absolute 
temperature, represents the heat transferred 
during the process (Fig. 4-8). 

Although the mathematical treatment of 
entropy is not required in the study of refrigera- 
tion and is beyond the scope of this book, it is 
important to note that according to Equation 
4-2 the entropy changes only when heat is 
transferred during the process. If there is no 
heat energy transfer, there is no change in the 
entropy. The heat energy transfer may occur 
either to or from an external source or sink or it 

* The average absolute temperature is not merely 
the mean of the initial and final temperatures of the 
process, but is the average of all of the absolute 
temperatures through which the process passes. 



may take place entirely within the fluid itself 
as a result of internal friction. However, the 
entropy of a fluid is not affected by external 
work done either by or on the fluid. Thus in a 
frictionless (occurring without either internal 
or external friction), adiabatic (no heat transfer 
to or from an external body) compression, as in 
the ideal compression of the refrigerant vapor 
in a refrigeration compressor, the entropy of the 
fluid will remain the same or constant. 
4-20. Vapor Tables. It has been stated 
previously that a vapor does not approach the 
condition of an ideal gas because of the inter- 
molecular forces which exist between the 
molecules of the vapor. Therefore, internal 
friction is present whenever a vapor undergoes a 
change of condition so that the various proper- 
ties of a vapor at the different conditions cannot 
be determined by applying the laws of ideal 



The properties of vapors at various conditions 
have been determined by experiment for all 
common vapors and these data are published in 
the form of tables. Separate tables are used for 
saturated and superheated vapors. 
4-21. Saturated Vapor Tables. Saturated 
vapor tables (Fig. 4-9) deal only with saturated 
liquids and vapors, and usually give values for 
the following properties: (1) temperature, (2) 
pressure, (3) specific volume, (4) enthalpy 
(specific), and (5) entropy (specific). Normally, 
the temperature in degrees Fahrenheit is listed 
in the extreme left-hand column. The pressure 
is given in the second and third columns, 
followed by the specific volume in cubic feet 
for both the liquid and the vapor in the fourth 
and fifth columns, respectively. Some tables 
list the density in addition to or in place of the 



SATURATED AND SUPERHEATED VAPORS 
Properties of Saturated Steam 



53 





Absolute Pressure 


Specific Volume 


Enthalpy 


Entropy 


Temp., 






Sat. 




Sat. 


Sat. 




Sat. 


Sat. 




Sat. 


"F, 

t 

(1) 


Psi, 

P 
(2) 


In. Hg, 

P 
(3) 


liquid, 
(4) 


Evap., 
(5) 


vapor, 

Vg 

(6) 


liquid, 
h, 
(7)« 


Evap., 

hfg 

(8) 


vapor, 

A. 

(9) 


liquid, 

s, 

(10) 


Evap., 

St, 
(11) 


vapor, 

s. 

(12) 


200 


11.526 


23.467 


0.01663 


33.62 


33.64 


167.99 


977.9 


1145.9 


0.2938 


1.4824 


1.7762 


202 


12.011 


24.455 


0.01665 


32.35 


32.37 


170.00 


976.6 


1146.6 


0.2969 


1.4760 


1.7729 


204 


12.512 


25.475 


0.01666 


31.14 


31.15 


172.02 


975.4 


1147.4 


0.2999 


1.4697 


1.7696 


206 


13.031 


26.531 


0.01667 


29.97 


29.99 


174.03 


974.2 


1148.2 


0.3029 


1.4634 


1,7663 


208 


13.568 


27.625 


0.01669 


28.86 


28.88 


176.04 


972.9 


1148.9 


0.3059 


1.4571 


1.7630 


210 


14.123 


28.755 


0.01670 


27.80 


27.82 


178.05 


971.6 


1149.7 


0.3090 


1.4508 


1.7598 


212 


14.696 


29.922 


0.01672 


26.78 


26.80 


180.07 


970.3 


1150.4 


0.3120 


1.4446 


1.7566 


214 


15.289 


31.129 


0.01673 


25.81 


25.83 


182.08 


969.0 


1151.1 


0.3149 


1.4385 


1.7534 


216 


15.901 


32.375 


0.01674 


24.88 


24.90 


184.10 


967.8 


1151.9 


0.3179 


1.4323 


1.7502 


218 


16.533 


33.662 


0.01676 


23.99 


24.01 


186.11 


966.5 


1152.6 


0.3209 


1.4262 


1.7471 


220 


17.186 


34.992 


0.01677 


23.13 


23.15 


188.13 


965.2 


1153.4 


0.3239 


1.4201 


1.7440 


222 


17.861 


36.365 


0.01679 


22.31 


22.33 


190.15 


963.9 


1154.1 


0.3268 


1.4141 


1.7409 


224 


18.557 


37.782 


0.01680 


21.53 


21.55 


192.17 


962.6 


1154.8 


0.3298 


1.4080 


1.7378 


226 


19.275 
C2O.O160 

20.780 


39.244 


0.01682 


20.78 


20.79 


194.18 


961.3 


1155.5 


0.3328 


1.4020 


1.7348 


228 


40.753 


0.01683 


20.06 


20.07 


196.20 


960.1 


1156.3 


0.3357 


1.3961 


1.7318 


230 


42.308 


0.01684 


19.365 


19.382 


198.23 


958.8 


1157.0 


0.3387 


1.3901 


1.7288 


240 


24.969 


50.837 


0.01692 


16.306 


16.323 


208.34 


952.2 


1160.5 


0.3531 


1.3609 


1.7140 


250 


29.825 


60.725 


0.01700 


13.804 


13.821 


218.48 


945.5 


1164.0 


0.3675 


1.3323 


1.6998 


260 


35.429 


72.134 


0.01709 


11.746 


11.763 


228.64 


938.7 


1167.3 


0.3817 


1.3043 


1.6860 


270 


41.858 


85.225 


0.01717 


10.044 


10.061 


238.841 931.8 


1170.6 


0.3958 


1.2769 


1.6727 



Fig. 4-9. Excerpt from typical saturated vapor table. 
Reproduced from Thermodynamic Properties of Steam by Keenan and Keyes, published by John Wiley and 
Sons, 1936, with permission. 



specific volume. If the density only is given 
and the specific volume is wanted, the specific 
volume is determined by dividing the density 
into one. Likewise, when the specific volume is 
given and the density is wanted, the density is 
found by dividing the specific volume into one 
(Section 3-4). 

Three values for enthalpy h are usually 
given in the saturated vapor tables: (1) the 
enthalpy of the liquid (h t ), which is the heat 
required to raise the temperature of the liquid 
from the temperature at the assumed zero point 
of enthalpy to the saturation temperature corre- 
sponding to the pressure of the liquid; (2) the 
enthalpy of vaporization (h fg ), which is the 



latent heat of vaporization at the pressure and 
temperature indicated; and (3) the enthalpy of 
the vapor (h g ), which is the sum of the enthalpy 
of the liquid (h f ) and the enthalpy of vapor- 
ization (h fa ). For example, the enthalpy of the 
liquid (h f ) for water at 212° F under atmospheric 
pressure is 180 Btu (1 x 1 x 180), whereas the 
enthalpy of the saturated water vapor at 212° F 
under atmospheric pressure is 1050 Btu, which 
is the sum of the enthalpy of the liquid (180 Btu) 
and the enthalpy of vaporization (970 Btu). 

Two values of entropy are usually given: 
s f , the entropy of the liquid and s g , the entropy 
of the vapor, the difference between the two 
being the change in entropy during vaporization. 



54 PRINCIPLES OF REFRIGERATION 



Dichlorodifluoromethane (Refrigerant-12) 
Properties of Superheated Vapor 





Abs. Pressure 36 lb/in. s 


Abs. Pressure 38 lb/in.* 


Abs. Pressure 40 lb/in. 2 


Abs. Pressure 42 lb/in. 2 


Temp. 


Gage Pressure 21.3 lb/in. 8 


Gage Pressure 23.3 lb/in. 8 


Gage Pressure 25.3 lb/in. 8 


Gage Pressure 27.3 lb/in. 8 


°F 


(Sat. Temp. 20.4° F) 


(Sat. Temp. 23.2° F) 


(Sat. Temp. 25.9° F) 


(Sat. Temp. 28.5° F) 


/ 


V 


H 


S 


V 


H 


S 


V 


H 


S 


V 


H 


S 


(at 
sat'n) 


{1.113) {80.54) {0.16947) 


{1.058) {80.86) {0.16931) 


{1.009) {81.16) (0.16914) 


(0,963) (81.44) (0.16897) 


30 


1.140 


81.90 


0.17227 


1.076 


81.82 


0.17126 


1.019 


81.76 


0.17030 


0.967 


81.65 


0.16939 


40 


1.168 

1.196 
1.223 


83.35 

84.81 
86.27 


0.17518 

0.17806 
0.18089 


1.103 

1.129 
1.156 


83.27 

84.72 
86.19 


0.17418 

0.17706 
0.17991 


1.044 


83.20 


0.17322 


0.991 

1.016 
1.040 


83.10 

84.56 
86.03 


0.17231 


50 


1.070 


84.65 


0.17612 


0.17521 


60 


1.095 


86.11 


0.17896 


0.17806 


70 


1.250 


87.74 


0.18369 


1.182 


87.67 


0.18272 


1.120 


87.60 


0.18178 


1.063 


87.51 


0.18086 


80 


1.278 


89.22 


0.18647 


1.208 


89.16 


0.18551 


1.144 


89.09 


0.18455 


1.087 


89.00 


0.18365 


90 


1.305 


90.71 


0.18921 


1.234 


90.66 


0.18826 


1.169 


90.58 


0.18731 


1.110 


90.50 


0.18640 


100 


1.332 


92.22 


0.19193 


1.260 


92.17 


0.19096 


1.194 


92.09 


0.19004 


1.134 


92.01 


0.18913 


110 


1.359 


93.75 


0.19462 


1.285 


93.69 


0.19365 


1.218 


93.62 


0.19272 


1.158 


93.54 


0.19184 


120 


1.386 


95.28 


0.19729 


1.310 


95.22 


0.19631 


1.242 


95.15 


0.19538 


1.181 


95.09 


0.19451 


130 


1.412 


96.82 


0.19991 


1.336 


96.76 


0.19895 


1.267 


96.70 


0.19803 


1.204 


96.64 


0.19714 


140 


1.439 


98.37 


0.20254 


1.361 


98.32 


0.20157 


1.291 


98.26 


0.20066 


1.227 


98.20 


0.19979 


ISO 


1.465 


99.93 


0.20512 


1.387 


99.89 


0.20416 


1.315 


99.83 


0.20325 


1.250 


99.77 


0.20237 


160 


1.492 


101.51 


0.20770 


1.412 


101.47 


0.20673 


1.340 


101.42 


0.20583 


1.274 


101.36 


0.20496 


170 


1.518 


103.11 


0.21024 


1.437 


103.07 


0.20929 


1.364 


103.02 


0.20838 


1.297 


102.96 


0.20751 


180 


1.545 


104.72 


0.21278 


1.462 


104.67 


0.21183 


1.388 


104.63 


0.21092 


1.320 


104.57 


0.21005 


190 


1.571 


106.34 


0.21528 


1.487 


106.29 


0.21433 


1.412 


106.25 


0.21343 


1.343 


106.19 


0.21256 


200 


1.597 


107.97 


0.21778 


1.512 


107.93 


0.21681 


1.435 


107.88 


0.21592 


1.365 


107.82 


0.21505 


210 


1.623 


109.61 


0.22024 


1.537 


109.57 


0.21928 


1.459 


109.52 


0.21840 


1.388 


109.47 


0.21754 


220 


1.650 


111.27 


0.22270 


1.562 


111.22 


0.22176 


1.482 


111.17 


0.22085 


1.411 


111.12 


0.22000 


230 


1.676 


112.94 


0.22513 


1.587 


112.89 


0.22419 


1.506 


112.84 


0.22329 


1.434 


112.80 


0.22244 


240 


1.702 


114.62 


0.22756 


1.612 


114.58 


0.22662 


1.530 


114.52 


0.22572 


1.457 


114.49 


0.22486 


250 


1.728 


116.31 


0.22996 


1.637 


116.28 


0.22903 


1.554 


116.21 


0.22813 


1.480 


116.19 


0.22728 


260 


1.754 


118.02 


0.23235 


1.662 


117.99 


0.23142 


1.577 


117.92 


0.23052 


1.502 


117.90 


0.22967 


270 


1.780 


119.74 


0.23472 


1.687 


119.71 


0.23379 


1.601 


119.65 


0.23289 


1.524 


119.62 


0.23204 


280 


1.807 


121.47 


0.23708 


1.712 


121.45 


0.23616 


1.625 


121.40 


0.23526 


1.547 


121.36 


0.23441 


290 


1.833 


123.22 


0.23942 


1.737 


123.20 


0.23850 


1.649 


123.15 


0.23760 


1.570 


123.11 


0.23675 


300 








1.762 


124.95 


0.24083 


1.673 


124.92 


0.23994 


1.592 


124.87 


0.23909 









Fig. 4-10. Excerpt from typical superheated vapor table. 
Copyright by E. I. du Pont de Nemours and Co., Inc. Reprinted by permission. 



It has been stated previously that the con- 
dition of a gas or a vapor can be determined 
when any two of its properties are known. How- 
ever, for a saturated liquid or vapor at any one 
pressure, there is only one temperature that the 
fluid can have and still satisfy the conditions of 
saturation. This is true also for the other 
properties of a saturated liquid or vapor. 
Therefore, if any one property of a saturated 
liquid or vapor is known, the value of the other 
properties can be read directly from the satu- 
rated vapor table. For instance, assume that 
the pressure of one pound of dry saturated 
steam is 20 psia. By locating 20 psia (encircled) 
in the second column of the abbreviated table in 



Fig. 4-9 and reading across the table, the values 
(set off by the heavy lines) for all the other 
properties of the vapor at this condition can be 
obtained. 

4-22. Superheated Vapor Tables. A super- 
heated vapor table deals with the properties of 
a superheated vapor rather than those of a 
saturated vapor, and the arrangement of a 
superheated vapor table is somewhat different 
from that of a saturated vapor table. One 
common form of the superheated vapor table is 
illustrated in Fig. 4-10. 

Before examining the superheated vapor 
table, it is important to take note of one 
significant difference between a saturated and a 



SATURATED AND SUPERHEATED VAPORS 55 



superheated vapor. Whereas, for a saturated 
vapor at any one pressure there is only one 
temperature which will satisfy the conditions 
of saturation, a superheated vapor may have 
any temperature above the saturation tempera- 
ture Corresponding to its pressure. The specific 
volume, enthalpy, and entropy of a superheated 
vapor at any one pressure will vary with the 
temperature. This does not mean that the 
properties of a superheated vapor are entirely 
independent of the pressure of the vapor but 
only that the properties of the superheated 
vapor at any one pressure will vary with the 
temperature. As a matter of fact, superheated 
vapor tables are based on the pressure of 
the vapor, and before the properties of a super- 
heated vapor can be determined from a table, the 
pressure of the vapor or one of the properties 
of the vapor at saturation must be known. 
When one of the properties of the vapor at 
saturation is known, the pressure of the vapor 
can be found by consulting a saturated vapor 
table. 

In addition to the properties of the super- 
heated vapor at various temperatures above the 
saturation temperature corresponding to the 
pressure, superheated vapor tables usually list 
some or all of the properties of the vapor at the 
saturation temperature. For example, in Fig. 
4-10, the absolute and gage pressures, along 
with the saturation temperature corresponding 
to these pressures, are given at the head of 
the table. The first readings in the body of the 
table (italicized) lists the specific volume, the 
enthalpy, and the entropy of the vapor at 
saturation. The specific volume, enthalpy, and 
entropy of the superheated vapor at various 
temperatures above the saturation temperature 
make up the body of the table. Notice that the 
temperature of the superheated vapor, given in 
the extreme left-hand column, is listed in 10° F 
increments. 

Example 4-1. One pound of superheated 

Refrigerant-12 vapor is at a temperature of 

50° F and its pressure is 40 psia. From the 

abbreviated table in Fig. 4-10, determine: 

(a) The temperature, volume, enthalpy, and 

entropy of the vapor at saturation 
(fc) The volume, enthalpy, and entropy of the 

vapor at the superheated condition 
(c) The degree of superheat of the vapor in 
degrees Fahrenheit 



(d) The amount of superheat in the vapor in 
Btu 

(e) The change in the volume during the 
superheating 

(/) The change in entropy during the super- 
heating 



Solution 




(a) From the head of 




the table, the satura- 




tion temperature 




corresponding to 




40 psia 


= 25.9° F 


From the body of 




the table (first 




reading, itali- 




cized),thespecific 




volume of the 




vapor at satura- 




tion 


= 1.009 cu ft/lb 


The enthalpy of the 




vapor at satura- 




tion 


= 81.16 Btu/lb 


The entropy of the 




vapor at satura- 




tion 


= 0.16914 Btu/lb/° R 


(b) From the body 




of the table, the 




properties of the 




vapor superheated 




to 50° F (offset by 




heavy lines in Fig. 




4-10) the specific 




volume 


= 1.070 cu ft/lb 


The enthalpy 


= 84.65 Btu/lb 


The entropy 


= 0.17612 Btu/lb/ R 


(c) The superheated 




temperature 


= 50.0° F 


The temperature at 




saturation 


= 25.9° F 


The degree of super- 




heat of the vapor 




in degrees Fahr- 




enheit 


= 24.1°F 


(rf)The enthalpy of 




the superheated 




vapor 


= 84.65 Btu/lb 


The enthalpy of the 




vapor at satura- 




tion 


= 81.16 Btu/lb 


The amount of 




superheat in the 




vapor in Btu 


= 3.49 Btu/lb 



56 PRINCIPLES OF REFRIGERATION 

(e) The entropy of the 

superheated vapor 

The entropy of the 
vapor at satura- 
tion 

The change in 
entropy during 
the superheating 



= 0.17612 Btu/lb/°R 



= 0.16914 Btu/lb/° R 



= 0.00698 Btu/lb/° R 



(/) The volume of the 
superheated vapor 
The volume of the 
vapor at satura- 
tion 
The change in vol- 
ume during the 
superheating 



= 1.070 cu ft/lb 
= 1.009 cu ft/lb 
= 0.061 cu ft/lb 



5 

Psych rometric 
Properties of Air 



5-1. Composition of Air. Air is a mecha- 
nical mixture of gases and water vapor. Dry 
air (air without water vapor) is composed 
chiefly of nitrogen (approximately 78% by 
volume) and oxygen (approximately 21 %), the 
remaining 1 % being made up of carbon dioxide 
and minute quantities of other gases, such as 
hydrogen, helium, neon, argon, etc. With 
regard to these dry air components, the com- 
position of the air is practically the same 
everywhere. On the other hand, the amount of 
water vapor in the air varies greatly with the 
particular locality and with the weather con- 
ditions. Since the water vapor in the air results 
primarily from the evaporation of water from 
the surface of various bodies of water, atmos- 
pheric humidity (water vapor content) is 
greatest in regions located near large bodies of 
water and is less in the more arid regions. 

Since all air in the natural state contains a 
certain amount of water vapor, no such thing 
as "dry air" actually exists. Nevertheless, the 
concept of "dry air" is a very useful one in that 
it greatly simplifies psychrometric calculations. 
Hereafter in this book the term "dry air" is 
used to denote air without water vapor, whereas 
the term "air" is used to mean the natural 
mixture of "dry air" and water vapor. 
5-2. Air Quantities. Air quantities may be 
stated either in units of volume (cubic feet) or in 
units of weight (pounds) so that the need for 
converting air quantities from one unit of 
measure to the other occurs frequently. 



The volume occupied by any given weight of 
air depends upon the pressure and temperature 
of the air, and varies inversely with the baro- 
metric pressure and directly with the absolute 
temperature. Air very nearly approaches the 
condition of an ideal gas and will follow the gas 
laws with sufficient accuracy for all practical 
purposes. Therefore, the volume occupied by 
any given weight of air at any given pressure 
and temperature can be determined by applying 
Equation 3-10. 

Example 5-1. Determine the volume oc- 
cupied by 1 lb of air having a temperature of 
70° F at standard sea level pressure (14.7 psia). 

Solution. Rearranging y _ M x R x T 
and applying Equation P 

3-10, 1 x 53.3 x 

(70 + 460) 
14.7 x 144 
= 13.34 cu ft 

Example 5-2. Determine the volume of the 
air in Example 5-1 if the barometric pressure is 
12.6 psia. 



Solution. Applying 
Equation 3-10, 



V = 



1 x 53.3 x 
(70 + 460) 



12.6 x 144 
= 15.57 cu ft 

Example 5-3. Determine the volume of the 
air in Example 5-1 if the temperature of the air 
is 100° F. 



Solution. Applying 
Equation 3-10, 



K = 



1 x 53.3 x 
(100 + 460) 



14.7 x 144 
= 14.10 cu ft 

The relationship between the volume and the 
weight of a given quantity of air at any condition 
is expressed by the following equations: 

V = M x v (5-1) 



V 
M = - 

v 



(5-2) 



where M = the weight of air in pounds 

V = the volume of M pounds of air in 

cubic feet 
v = the specific volume of the air in 

cubic feet per pound 

Example 5-4. Air at a temperature of 
95° F is circulated over a cooling coil at the 
rate of 2000 cu ft/min (cfm). If the specific 



57 



58 PRINCIPLES OF REFRIGERATION 

volume of the air is 14.38 cu ft/lb, determine 
the weight of air passing over the coil in pounds 
per hour. 

Solution. Applying M _ 

Equation 5-2, the weight 
of air passing over the 
cooling coil 

Multiplying by 60 min 



2000 



14.38 
= 139.2 lb/min 

M - 139.1 x 60 
= 8346 lb/hr 

5-3. Standard Air. Because of the difference 
in the volume of any given weight of air at 
various temperatures and pressures, an air 
standard has been established for use in the 
rating of air handling equipment so that all 
equipment is rated at equal conditions. Dry air 
having a specific volume of 1 3.34 cu ft per pound 
or a density of 0.07496 (0.075) lb per cu ft 
(1/13.34) is defined as standard air. Air at a 
temperature of 70° F and at standard sea level 
pressure has this specific volume and density 
(see Example 5-1). 

A given volume of air at any condition can be 
converted to an equivalent volume of standard 
air by applying the following equation : 






(5-3) 



where V s = the equivalent volume of standard 
air 
V a = the actual volume of the air at any 

given condition 
v a =* the specific volume of the air at the 

given condition 
v, = the specific volume of standard air 
(13.34 cu ft/lb) 

Example 5-5. For the air in Example 5-4, 
determine the equivalent volume of standard 
air. 

Solution. Applying 
Equation 5-3, the equiv- 
alent volume of standard 
airK, 



_ 2000 x 14.38 

13.34 
= 2155 cfm 



5-4. Dalton's Law of Partial Pressure. 

Dalton's law of partial pressures states in effect 
that in any mechanical mixture of gases and 
vapors (those which do not combine chemically) : 
(1) each gas or vapor in the mixture exerts an 
individual partial pressure which is equal to the 
pressure that the gas would exert if it occupied 
the space alone and (2) the total pressure of the 
gaseous mixture is equal to the sum of the 



partial pressures exerted by the individual 
gases or vapors. 

Air, being a mechanical mixture of gases and 
water vapor, obeys Dalton's law. Therefore, 
the total barometric pressure is always equal to 
the sum of the partial pressures of the dry gases 
and the partial pressure of the water vapor. 
Since psychrometry is the study of the properties 
of air as affected by the water vapor content, 
the individual partial pressures exerted by the 
dry gases are unimportant and, for all practical 
-purposes, the total barometric pressure may be 
considered to be the sum of only two pressures: 
(1) the partial pressure exerted by the dry gases 
and (2) the partial pressure exerted by the water 
vapor. 

5-5. Dew Point Temperature. It is im- 
portant to recognize that the water vapor in 
the air is actually steam at low pressure and that 
this low pressure steam, like high pressure steam 
will be in a saturated condition when its 
temperature is the saturation temperature corre- 
sponding to its pressure. Since all of the com- 
ponents in a gaseous mixture are at the same 
temperature, it follows that when air is at any 
temperature above the saturation temperature 
corresponding to the partial pressure exerted by 
the water vapor the water vapor in the air will 
be superheated. On the other hand, when air 
is at a temperature equal to the saturation 
temperature corresponding to the partial 
pressure of the water vapor, the water vapor in 
the air is saturated and the air is said to be 
saturated (actually it is only the water vapor 
which is saturated). The temperature at which 
the water vapor in the air is saturated is known 
as the dew point temperature of the air. Ob- 
viously, then, the dew point temperature of the 
air is always the saturation temperature corre- 
sponding to the partial pressure exerted by the 
water vapor. Hence, when the partial pressure 
exerted by the water vapor is known, the dew 
point temperature of the air can be determined 
from the steam tables. Likewise, when the dew 
point temperature of the air is known, the 
partial pressure exerted by the water vapor can 
be determined from the steam tables. 

Example 5-6. Assume that a certain quan- 
tity of air has a temperature of 80° F and that 
the partial pressure exerted by the water vapor 
in the air is 0.17811 psia. Determine the dew 
point temperature of the air. 



PSYCHROMETRIC PROPERTIES OF AIR 59 



Solution. From Table 4-1, the saturation 
temperature of steam corresponding to a pres- 
sure of 0.17811 psia is 50° F. Therefore, 50° F 
is the dew point temperature of the air. 

Example 5-7. A certain quantity of air has 
a temperature of 80° F and a dew point tem- 
perature of 40° F. Determine the partial pres- 
sure exerted by the water vapor in the air. 

Solution. From Table 4-1, the saturation 
pressure corresponding to 40° F is 0.12170 psia 
and therefore 0.12170 psia is the partial pressure 
exerted by the water vapor. 

It has been shown (Section 4-5) that the 
pressure exerted by any vapor is directly 
proportional to the density (weight per unit 
volume) of the vapor. Since the dew point 
temperature of the air depends only on the 
partial pressure exerted by the water vapor, it 
follows that, for any given volume of air, the 
dew point temperature of the air depends only 
upon the weight of water vapor in the air. As 
long as the weight of water vapor in the air 
remains unchanged, the dew point temperature 
of the air will also remain unchanged. If the 
amount of water vapor in the air is increased or 
decreased, the dew point temperature of the air 
will also be increased or decreased, respectively. 
Increasing the amount of water vapor in the air 
will increase the pressure exerted by the water 
vapor and raise the dew point temperature. 
Likewise, reducing the amount of water vapor 
in the air will reduce the pressure of the water 
vapor and lower the dew point temperature. 
5-6. Maximum Water Vapor Content. 
The maximum amount of water vapor that can 
be contained in any given volume of air depends 
only upon the temperature of the air. Since the 
amount of water vapor in the air determines the 
partial pressure exerted by the water vapor, it 
is evident that the air will contain the maximum 
amount of water vapor when the water vapor 
in the air exerts the maximum possible pressure. 
Since the maximum pressure that can be 
exerted by any vapor is the saturation pressure 
corresponding to its temperature, the air will 
contain the maximum weight of water vapor 
when the pressure exerted by the water vapor 
is equal to the saturation pressure corresponding 
to the temperature of the air. At this condition 
the temperature of the air and the dew point 
temperature will be one and the same and the air 



will be saturated. It is important to notice that 
the higher the temperature of the air, the higher 
is the maximum possible vapor pressure and the 
greater is the maximum possible water vapor 
content. 

5-7. Absolute Humidity. The water vapor 
in the air is called humidity. The absolute 
humidity of the air at any given condition is 
denned as the actual weight of water vapor 
contained in 1 cu ft of air at that condition. 
Since the weight of water vapor contained in the 
air is relatively small, it is often measured in 
grains rather than in pounds (7000 grains equal 
lib). 

5-8. The Psychrometric Tables. It was 
shown in Section 5-5 that the actual weight of 
water vapor contained in a unit volume of air is 
solely a function of the dew point temperature 
of the air. Because of this fixed relationship 
between the dew point temperature and the 
absolute humidity of the air, when the value 
of one is known, the value of the other can be 
readily computed. 

The absolute humidity of air at various dew 
point temperatures is listed in Tables 5-1 and 
5-2. The dew point temperatures are listed in 
column (1) of the tables, and the absolute 
humidity corresponding to each of the dew 
point temperatures is given in columns (4) and 
(5). The values given in column (4) are in 
pounds of water vapor per cubic foot of air, 
whereas the values given in column (5) are in 
grains of water vapor per cubic foot of air. 
Too, the partial pressure (saturation pressure) 
of the vapor corresponding to each dew point 
temperature is given in inches of mercury in 
column (2) and in pounds per square inch in 
column (3). 

5-9. Relative Humidity. Relative humidity 
(RH), expressed in percent, is the ratio of the 
actual weight of water vapor per cubic foot of 
air relative to the weight of water vapor con- 
tained in a cubic foot of saturated air at the 
same temperature, viz: 

Actual weight of 
water vapor per 

cubic foot of air 

Relative humidity = ... . . . -z — x 100 

J Weight of water 

vapor in 1 cu ft of 

saturated air at the 

same temperature 



60 PRINCIPLES OF REFRIGERATION 



For instance, if air at a certain temperature 
contains only half as much water vapor per 
cubic foot of air as the air could contain at that 
temperature if it were saturated, the relative 
humidity of the air is 50%. The relative 
humidity of saturated air, of course, is 100%. 

Example 5-8. Air at a temperature of 
80° F has a dew point temperature of 50° F. 
Determine the relative humidity. 

Solution. From 
Table 5-2, absolute 
humidity correspond- 
ing to dew point tem- 
perature of 50° F 

Absolute humidity 
of saturated air at 
80° F 

Applying Equation _ 
5-4, the relative humid- — 11,04 
ity of the air = 37.1 % 



= 4.106 grains/cu ft 



11.04 grains/cu ft 

4.106 1/MV 
x 100 



Example 5-9. Determine the relative hu- 
midity of the air in Example 5-8, if the air is 
cooled to 60° F. (Note : the dew point tempera- 
ture of the air does not change because the 
moisture content does not change.) 

Solution. From Table 
5-2, absolute humidity 
corresponding to dew 
point of 50° F 

Absolute humidity of 
saturated air at 60° F 

Applying Equation 5-4, 
the relative humidity of 
the air 



= 4.106 grains/cu ft 

= 5.795 grains/cu ft 
_ 4.106 
5.795 
= 70.8% 



5-10. Specific Humidity. The specific humid- 
ity is the actual weight of water vapor mixed 
with 1 lb of dry air and is usually stated in 
grains per pound, that is, grains of water vapor 
per pound of dry air. For any given barometric 
pressure, the specific humidity is a function of 
the dew point temperature alone. The specific 
humidity of air at various dew point tempera- 
tuies is listed in Columns 6 and 7 of Tables 5-1 
and 5-2. In Column 6, the specific humidity is 
given in pounds of water vapor per pound of dry 
air, whereas in Column 7 the specific humidity 
is given in grains of water vapor per pound of 
dry air. Since the specific humidity correspond- 
ing to any given dew point temperature varies 
with the total barometric pressure, the values 



given in Tables 5-1 and 5-2 apply only to air at 
standard barometric pressure. 

The specific humidity of the air at any given 
dew point temperature increases as the total 
barometric pressure decreases and decreases as 
the total barometric pressure increases. The 
reason for this is easily explained. It has been 
shown (Examples 5-1 and 5-3) that the volume 
occupied by 1 lb of air increases as the total 
barometric pressure decreases. Since the density 
of a vapor varies inversely with the volume, it 
follows that the weight of water vapor required 
to produce a given vapor density and vapor 
pressure increases as the volume of the air 
increases. Likewise, as the volume occupied by 
1 lb of air diminishes, the weight of water vapor 
required to produce a certain vapor density and 
vapor pressure also diminishes. 
5-11. Percentage Humidity. Percentage 
humidity is defined as the ratio of the actual 
weight of water vapor in the air per pound of dry 
air to the weight of water vapor required to 
saturate completely 1 lb of dry air at the same 
temperature. Percentage humidity, like relative 
humidity, is given in percent. Notice, however, 
that percentage humidity is associated with the 
weight of water vapor per unit weight of air, 
whereas relative humidity is associated with the 
weight of water vapor per unit volume of air. 
For this reason the percentage humidity varies 
with the total barometric pressure, whereas 
relative humidity does not. 

Example 5-10. Air at standard sea level 
pressure has a temperature of 80° F and a dew 
point temperature of 50° F. Determine the 
specific humidity and percentage humidity of 
the air. 



Solution. From Table 
5-2, the specific humidity 
of the air in grains per 
pound corresponding to 
a 50° F dew point tem- 
perature (Column 7) 

Specific humidity of 
saturated air at 80° F 
(Column 7) 

Percentage humidity 



53.38 grains/lb 



155.50 grains/lb 
53.38 



155.50 

34.3% 



x 100 



Note. Compare this value with the relative 
humidity obtained in Example 5-8. 



PSYCHROMETRIC PROPERTIES OF AIR 61 



5-12. Dry Bulb and Wet Bulb Tempera- 
tures. The dry bulb (DB) temperature of the 
air is the temperature as measured by an ordinary 
dry bulb thermometer. When measuring the dry 
bulb temperature of the air, the bulb of the ther- 
mometer should be shaded to reduce the effects 
of direct radiation. 

The wet bulb (WB) temperature of the air is 
the temperature as measured by a wet bulb ther- 
mometer. A wet bulb thermometer is an ordi- 
nary thermometer whose bulb is enclosed in a 
wetted cloth sac or wick. To obtain an accurate 
reading with a wet bulb thermometer, the wick 
should be wetted with clean water at approxi- 
mately the dry bulb temperature of the air and 
the air velocity around the wick should be main- 
tained between 1000 and 2000 ft per minute. As 
a practical matter, this velocity can be simulated 
in still air by whirling the thermometer about on 
the end of a chain. An instrument especially 
designed for this purpose is the sling psychrom- 
eter (Fig. 5-1). The sling psychrometer is made 
up of two thermometers, one dry bulb and one 
wet bulb, mounted side by side in a protective 
case which is attached to a handle by a swivel 
connection so that the case can be easily rotated 
about the hand. After saturating the wick with 
clean water, the instrument is whirled rapidly in 
the air for approximately one minute, after which 
time readings can be taken from both the dry 
bulb and wet bulb thermometers. The process 
should be repeated several times to assure that 
the lowest possible wet bulb temperature has 
been recorded. 

Unless the air is 100 % saturated, in which case 
the dry bulb, wet bulb, and dew point tempera- 
tures of the air will be one and the same, the 
temperature recorded by a wet bulb thermom- 
eter will always be lower than the dry bulb 
temperature of the air. The amount by which 
the wet bulb temperature is reduced below the 
dry bulb temperature depends upon the relative 
humidity of the air and is called the wet bulb 
depression. 

Whereas a dry bulb thermometer, being un- 
affected by humidity, measures only the actual 
temperature of the air, a wet bulb thermometer, 
because of its wetted wick, is greatly influenced 
by the moisture in the air; thus a wet bulb tem- 
perature is in effect a measure of the relationship 
between the dry bulb temperature of the air and 
the moisture content of the air. In general, for 



Swivel 
connection 



Wet bulb 
thermometer 




.Dry bulb 
"thermometer 



Wetted wick 



Fig. 5-1. Sling psychrometer. 

any given dry bulb temperature, the lower the 
moisture content of the air, the lower is the wet 
bulb temperature. The reason for this is easily 
explained. 

When unsaturated air is brought into contact 
with water, water will evaporate into the air at a 
rate proportional to the difference in pressure 
between the vapor pressure of the water and the 
vapor pressure of the air. Hence, when a wet 
bulb thermometer is whirled rapidly about in 
unsaturated air, water will evaporate from the 
wick, thereby cooling the water remaining in the 
wick (and the thermometer bulb) to some tem- 
perature below the dry bulb temperature of the 
air. 

It is important to recognize the fact that the 
wet bulb temperature of the air is a measure of 
the relationship between the dry bulb and dew 
point temperatures of the air, and as such it 
provides a convenient means of determining the 
dew point temperature of the air when the dry 



62 PRINCIPLES OF REFRIGERATION 



bulb temperature is known. Too, it will be 
shown later that the wet bulb temperature is also 
an index of the total heat content of the air. 

In order to understand why the wet bulb tem- 
perature is a measure of the relationship between 
the dry bulb and dew point temperatures, a know- 
ledge of the theory of the wet bulb thermometer 
is required. When water evaporates from the 
wick of a wet bulb thermometer, heat must be 
supplied to furnish the latent heat of vaporiza- 
tion. Before the temperature of the water in the 
wick is reduced below the dry bulb temperature 
of the air, the source of the heat to vaporize the 
water is the water itself. Therefore, as water 
evaporates from the wick, the water remaining 
in the wick is cooled below the dry bulb tempera- 
ture of the air. When this occurs, a temperature 
differential is established and heat begins to flow 
from the air to the wick. Under this condition, a 
part of the vaporization heat is being supplied by 
the air while the other part is supplied by the 
water in the wick. As the temperature of the 
wick continues to decrease, the temperature 
difference between the air and the wick increases 
progressively so that more and more of the 
vaporization heat is supplied by the air and less 
and less is supplied by the water in the wick. 
When the temperature of the wick is reduced to 
the point where the temperature difference 
between the air and the "wick is such that the 
flow of heat from the air is sufficient to supply all 
of the vaporizing heat, the temperature of the 
wick will stabilize even though vaporization from 
the wick continues. The temperature at which 
the wick stabilizes is called the temperature of 
adiabatic saturation and is the wet bulb tempera- 
ture of the air. 

Through careful analysis of the foregoing, it 
can be seen that the wet bulb temperature 
depends upon both the dry bulb temperature and 
the amount of water vapor in the air. For 
example, the lower the relative humidity of the 
air, the greater is the rate of evaporation from 
the wick and the greater is the amount of heat 
required for vaporization. Obviously, the greater 
the need for heat, the greater is the wet bulb 
depression below the dry bulb temperature. Too, 
it follows also that the lower the dry bulb tem- 
perature, the lower the wet bulb temperature for 
any given wet bulb depression. 
5-13. The Heat Content or Enthalpy of 

Air. Air has both sensible and latent heat, 



and the total heat content of the air at any 
condition is the sum of the sensible and latent 
heat contained therein. 

The sensible heat of the air is a function of the 
dry bulb temperature. For any given dry bulb 
temperature, the sensible heat of the air is taken 
as the enthalpy of dry air at that temperature as 
calculated from 0° F. Air sensible heat at various 
temperatures is given in Btu per pound of dry air 
in Column 10 of Tables 5-1 and 5-2. With regard 
to Column 10, the temperatures listed in Column 
1 are used as dry bulb temperatures. 

Example 5-1 1. Using Table 5-2, determine 
the sensible heat in 10 lb of air at 80° F. 



Solution. From Table 
5-2, the sensible heat of 1 lb 
of air at 80° F 

For 10 lb of air, the sen- 
sible heat at 80° F 



= 19.19 Btu/lb 
= 10 x 19.19 
= 191 .9 Btu 



The quantity of sensible heat added or 
removed in heating or cooling a given weight of 
air through a given temperature range may be 
computed by applying Equation 2-8. The mean 
specific heat of air at constant pressure is 0.24 
Btu/lb. (Although the specific heat of any vapor 
or gas varies somewhat with the temperature 
range, the use of a mean specific heat value is 
sufficiently accurate for all practical purposes.) 

Example 5-12. Compute the quantity of 
sensible heat required to raise the temperature 
of 10 lb of air from 0° F to 80° F. 



Solution. Applying 
Equation 2-8, Q s 

Alternate Solution. From 
Table 5-2, the sensible heat 
of 1 lb of air at 80° F 

Sensible heat of 1 lb of 
air at 0° F 

For 1 lb of air, Q a 

For 10 lb of air, Q, 



10 x 0.24 

x (80 - 0) 
192 Btu 



= 19.19 Btu/lb 

= Btu/lb 
= 19.19 -0 
= 19.19 Btu/lb 
= 10 x 19.19 
= 191.9 Btu 



Since all the components of dry air are non- 
condensable at normal temperatures and pres- 
sures, for all practical purposes the only latent 
heat in the air is the latent heat of the water 
vapor in the air. Therefore, the amount of 



PSYCHROMETRIC PROPERTIES OF AIR 63 



latent heat in any given quantity of air depends 
upon the weight of water vapor in the air and 
upon the latent heat of vaporization of water 
corresponding to the saturation temperature of 
the water vapor. 

Since the saturation temperature of the water 
vapor is the dew point temperature of the air, 
the dew point temperature determines not only 
the weight of water vapor in the air but also the 
value of the latent heat of vaporization. Hence, 
the latent heat content of the air is a function of 
the dew point temperature alone. As long as the 
dew point temperature of the air remains un- 
changed, the latent heat content of the air also 
remains unchanged. 

The total heat content of water vapor at 
various temperatures as computed from 32° F is 
given in Btu per pound in Column 1 1 of Tables 
5-1 and 5-2. Although the values given in 
Column 11 include the sensible heat of the 
liquid above 32° F as well as the latent heat of 
vaporization at the given temperature, common 
practice is to treat the entire heat content of the 
water vapor as latent heat.* 

The latent heat content of any given quantity 
of air can be computed by multiplying the actual 
weight of water vapor in the air in pounds by the 
total heat of the water vapor as given in Column 
11 of Tables 5-1 and 5-2. 

Example 5-13. Compute the latent heat 
content of the air in Example 5-12, if the dew 
point temperature of the air is 50° F. 

Solution. From 
Table 5-2, the actual 
weight of water vapor 
per pound of dry air 
(specific humidity) at 
50° F DP (Column 6) = 0.007626 lb 

Total heat per 
pound of saturated 
water vapor at 50° F 
(Column 11) = 1081.7 Btu/lb 

* Although the total heat of the air at any con- 
dition is the sum of the sensible and latent heat 
contained therein, as a practical matter it is more 
convenient to consider the total enthalpy of the air 
as being the sum of the enthalpy of the dry air and 
the enthalpy of the water vapor mixed with the dry 
air. Since the amount of sensible heat is com- 
paratively small, the error which accrues from 
assuming all of the heat of the vapor to be latent 
heat is of no practical consequence. 



= 0.007626 x 1081.7 
= 8.25 Btu/lb 
= 10 x 8.25 
= 82.5 Btu 



Latent heat per 
pound of dry air at 
50° F DP 

For 10 lb of dry air, 
total latent heat 

Since the total heat of the air is the sum of the 
sensible and latent heat contained therein, the 
total heat of the air in Examples 5-12 and 5-13 
is the sum of the sensible heat of the dry air, as 
computed in Example 5-12, and the latent heat 
of the water vapor mixed with the dry air, as 
computed in Example 5-13, viz: 

Sensible heat of 10 lb of dry 
air at 70° F, from Example 
5-12 

Latent heat of water vapor 
mixed with 10 lb of dry air at 
50° F DP, from Example 5-13 

Total heat of 10 lb of air at 
70° F DB and 50° F DP* 



- 191.9 Btu 



82.5 Btu 
191.9 + 82.5 
274.4 Btu 



5-14. Wet Bulb Temperature as a Measure 
of Total Heat. It has been shown in pre- 
ceding sections that the sensible heat of the air 
(the heat content of the dry air) is a function of 
the dry bulb temperature and that the latent heat 
of the air (the heat content of the water vapor 
mixed with the dry air) is a function of the dew 
point temperature. Since, for any given com- 
bination of dry bulb and dew point temperatures, 
the wet bulb temperature of the air can have only 
one value, it is evident that the wet bulb tempera- 
ture is an index of the total heat content of the 
air. However, it is important to recognize that 
although there is only one wet bulb temperature 
that will satisfy any given combination of dry 
bulb and dew point temperatures, there are 
many combinations of dry bulb and dew point 
temperatures which will have the same wet bulb 
temperature (see Fig. 5-2). This means in effect 
that different samples of air having the same wet 

* The actual weight of air involved is slightly in 
excess of 10 lb, being 10 lb of dry air plus the weight 
of water vapor (0.007626 lb) mixed with the dry air. 
Too, since the temperature of the water vapor is the 
same as that of the dry air (70° F), the water vapor 
contains a certain amount of superheat (50° F to 
70° F) which is not included in the total heat. How- 
ever, since both of these values are very small, the 
error incurred by neglecting them has no practical 
significance. 



64 PRINCIPLES OF REFRIGERATION 



Temperature, 


°F 


Heat Content, Btu/lb 


Dry 


Dew 


Wet 








Bulb 


Point 


Bulb 


Sensible 


Latent 


Total 


60 


60 


60 


14.39 


11.98 


26.37 


65 


57 


60 


15.59 


10.78 


26.37 


70 


53.5 


60 


16.79 


9.58 


26.37 


75 


50 


60 


17.99 


8.38 


26.37 


80 


45.5 


60 


19.19 


7.18 


26.37 


85 


40.5 


60 


20.39 


5.98 


26.37 


90 


34.5 


60 


21.59 


4.78 


26.37 



Fig. 5-2 

bulb temperature have the same total heat, even 
though the ratio of sensible to latent heat may be 
different for the different samples. 

The values of total heat listed for various 
temperatures in Column 12 of Tables 5-1 and 
5-2 are for 1 lb of saturated air at the tempera- 
ture shown. However, since all samples of air, 
saturated or unsaturated, having the same wet 
bulb temperature have the same total heat, the 
values given in Column 12 will apply to any 
sample of air when the temperatures listed in 
Column 1 are used as wet bulb temperatures. 

Example 5-14. If 100 lb of air having an 
initial wet bulb temperature of 78° F are cooled 
to a final wet bulb temperature of 60° F, deter- 
mine the total heat removed from the air during 
the cooling process. 

Solution. From Table 
5-2, total heat of 1 lb of 
air corresponding to 78" F 
WB (Column 12) - 63.05 Btu/lb 

Total heat per pound of 
air at 60° F WB (Column 
12) 

Total heat removed per 
pound of air, Q t 



For 1001b of air, the 
total heat removed, Q t 



= 26.37 Btu/lb 
= 63.05 - 26.37 
= 36.68 Btu/lb 

- 100 x 36.68 
= 366.8 Btu 

5-15. Specific Volume of Air. It has already 
been shown that the volume occupied by a given 
weight of air depends upon the temperature of 
the air and upon the total barometric pressure. 
For standard sea level pressure, the volume of 
1 lb of dry air at various temperatures is listed 
in Column 8 of Tables 5-1 and 5-2. The volume 
of 1 lb of saturated air (1 lb of dry air and the 
water vapor to saturate it) is listed for various 
temperatures in Column 9. When the relative 



humidity of the air is known, the specific volume 
of partially saturated air at any condition can be 
computed by applying these values in the follow- 
ing equation: 

v a = v* + K»» - v d ) x %RH] (5-5) 

where v a = the specific volume of partially 

saturated air 
v d = the specific volume of dry air at the 

same temperature 
v s = the specific volume of saturated air 

at the same temperature 

Example 5-15. Compute the specific vol- 
ume of air at 95° F DB and 50% RH. 



Solution. From Table 
5-2, specific volume of 
dry air at 95° F (Column 
8) 

Specific volume of 
saturated air at 95° F 
(Column 9) 

Applying Equation 
5-5, v a 



= 1 3.97 cu ft/lb 



= 14.79 cu ft/lb 
= 13.97 + [(14.79 
- 13.97) x 0.5] 
= 14.38 cu ft/lb 

5-16. The Psychrometric Chart. Psychro- 
metric charts (Fig. 5-3) are graphical repre- 
sentations of psychrometric data such as those 
contained in Tables 5-1 and 5-2. The use of 
psychrometric charts permits graphical analysis 
of psychrometric data and thereby facilitates the 
solution of many practical problems dealing with 
air which would otherwise require tedious mathe- 
matical calculation. 

Basically, the psychrometric chart shows the 
relationship between four fundamental proper- 
ties of air: (1) dry bulb temperature, (2) dew 
point temperature, (3) wet bulb temperature, and 
(4) relative humidity. When any two of these 
four properties are known, the other two can be 
determined directly from the psychrometric chart 
without using mathematical calculations. 

The skeleton chart in Fig. 5-4 illustrates the 
general construction of the psychrometric chart 
which is based primarily upon the relationship 
that exists between the aforementioned four pro- 
perties. Notice that the lines of dry bulb tem- 
perature are vertical while the lines of dew point 
temperature are horizontal. The lines of wet 
bulb temperature run diagonally across the chart 
as do the lines of constant volume. The curved 



PSYCHROMETRIC PROPERTIES OF AIR 65 



lines are lines of constant relative humidity. The 
curved line bounding the chart on the left side is 
the line of 100% relative humidity and is called 
the saturation curve. Air at any condition such 
that its state can be identified by a point falling 
anywhere along the saturation curve is saturated 
air. Values for dry bulb, wet bulb, and dew 
point temperatures are read at the saturation 
curve. Values for dry bulb temperature are also 
given at the base of the chart. Notice that the 
dry bulb, wet bulb, and dew point temperatures 
for saturated air coincide. Values of specific 
volume and relative humidity are given along 
the lines of constant volume and relative humid- 
ity, respectively. Values of specific humidity and 
vapor pressure are given on the right and left 
margins of the chart. For any given air condi- 
tion, the specific humidity and vapor pressure 
corresponding to the dew point temperature can 
be determined by following the dew point tem- 
perature line to the specific humidity and vapor 
pressure scales. The total heat corresponding to 
any wet bulb temperature is found by following 
the wet bulb lines to the total heat scale above 
the saturation curve. The following example will 
illustrate the use of the psychrometric chart. 

Example 5-16. A certain quantity of air has 
a dry bulb temperature of 95° F and a wet bulb 
temperature of 77° F. From the psychrometric 
chart determine all of the following values: (1) 
dew point temperature, (2) specific humidity, 
(3) vapor pressure, (4) specific volume, (5) total 
heat, and (6) relative humidity. 

Solution. Using the two known properties of 
the air as coordinates the condition of the air 
can be established as a point on the chart. Once 
this point has been established, the other pro- 
perties of the air at this condition can be read 
directly from the chart as shown in Fig. 5-5, viz: 

Dew point temperature = 70° F 

Specific humidity =110 grains/lb 

Vapor pressure = 0.37 psia 

Specific volume = 14.33 cu ft/lb 

Total heat = 40.5 Btu/lb 

Relative humidity = 45 % 

Example 5-17. For the air in Example 5-16, 
determine: (a) the sensible heat per pound of 
air and (6) the latent heat per pound of air. 

Solution, 
(a) From Table 5-2, en- 
thalpy of 1 lb of dry air 
at 95° F DB, Q, = 22.80 Btu/lb 



(b) From Example 5-16, 
total heat per pound of 
air, Q t 

The latent heat per 
pound of air, & 



= 40.50 Btu/lb 
= Qt - Q, 

= 40.50 - 22.80 
= 17.7 Btu/lb 



Example 5-18. If the air in Example 5-16 is 
cooled to 75° F, determine: 

(a) The final dew point temperature 

(b) The final wet bulb temperature 

(c) The final relative humidity 

(d) The final total heat per pound 

Solution. Since the air is not cooled below the 
initial dew point temperature, no moisture is 
removed from the air. Therefore, the specific 
humidity, dew point temperature, and latent 
heat of the air remain unchanged. Hence, the 
initial dew point temperature and the new dry 
bulb temperature can be used as coordinates to 
locate the new condition of the air on the 
psychrometric chart (point B in Fig. 5-6). The 
following properties of the air at the new con- 
dition are taken from the psychrometric chart 
as indicated in Fig. 5-6: 

(a) Wet bulb temperature = 71.4° F 

(b) Relative humidity = 85 % 

(c) Total heat per pound = 35.69 Btu/lb 

Example 5-19. With respect to Fig. 5-6, in 
cooling the air from condition "A," as described 
in Example 5-16, to condition "B," as described 
in Example 5-18, compute: 

(a) The total heat removed per pound of air 

(b) The sensible heat removed per pound of 
air. 



Solution. 
(a) From Example 5-16, 
air total heat at A 

From Example 5-18, air 

total heat at B 

The total heat removed 

per pound of air 

in cooling from A 

toB 



= 40.50 Btu/lb 
= 35.69 Btu/lb 



= 40.50 - 35.69 
= 4.81 Btu/lb 



(b) Since there is no change in the latent heat of 
the air, the sensible heat removed per pound 
of air is equal to the total heat removed 
per pound of air. 

Example 5-20. Assume that the air in 
Example 5-16 is cooled to 40° F and determine: 

(a) The total heat removed per pound 

(b) The sensible heat removed per pound 

(c) The latent heat removed per pound. 



66 PRINCIPLES OF REFRIGERATION 



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70 PRINCIPLES OF REFRIGERATION 



Solution. Since the air is cooled below the 
dew point temperature, moisture will be con- 
densed out of the air and the air at the final 
condition will be saturated. Therefore, the dry 
bulb temperature, the dew point temperature, 
and the wet bulb temperature will coincide at 
40° F and the relative humidity of the air will be 
100%. On the psychrometric chart, the con- 
dition of the air falls on the saturation curve at 
40° F (point B in Fig. 5-7). 

(a) From the psychromet- 
ric chart, the total heat 
of the air at the initial 

condition (point A) = 40.50 Btu/lb 

The total heat of the air 

at the final condition 

(point B) =15.19 Btu/lb 

The total heat removed 40.50 - 15.19 

per pound = 25.31 Btu/lb 

(b) Applying Equation 2-8, 

the sensible heat re- = 1 x 0.24 
moved per pound of x (95 — 40) 

air =13.2 Btu/lb 

(c) The latent heat re- = Q t - Q, 
removed per pound = 25.31 — 13.20 
of air =12.11 Btu/lb 

PROBLEMS 

1. Determine the volume occupied by 1 lb of air 
having a temperature of 80° F at standard sea 
level pressure. Ans. 13.59 cu ft 

2. Compute the volume of the air in Problem 1 
if the barometric pressure is 13.5 psia. 

Ans. 14.80 cu ft 

3. Determine the volume of the air in Problem 1 
if the temperature of the air is 120° F. 

Ans. 14.60 cu ft 

4. Air at a temperature of 90° F is circulated 
over a cooling coil at the rate of 1000 cu ft per 



min (cfm). If the specific volume of the air is 
14.10 cu ft/lb, compute the weight of air passing 
over the coil in pounds per hour. 

Ans. 4255 lb/hr 

5. Compute the equivalent volume of standard 
air for the conditions of Problem 4. 

Ans. 946 cfm 

6. Compute the quantity of sensible heat re- 
quired to raise the temperature of 10 lb of air 
from a temperature of 35° F to a temperature of 
100° F. Ans. 156 Btu 

7. If 80 lb of air having an initial wet bulb tem- 
perature of 80° F are cooled to a final wet bulb 
temperature of 65° F, determine the total heat 
removed from the air during the cooling process. 

Ans. 1085.6 Btu 

8. A certain quantity of air has "a dry bulb 
temperature of 90° F and a wet bulb tempera- 
ture of 77° F. From the psychrometric chart 
determine all of the following values: 

(a) dew point temperature, (£>) specific humidity, 
(c) vapor pressure, (d) specific volume, (e) total 
heat, and (/) relative humidity. 
Ans. (a) 72.3° F; (6) 119.5 gpp; (c) 0.395 psia; 
(a") 14.23 cu ft/lb; (e) 40.5 Btu/lb; (/) 58% 

9. Assume that the air in Problem 8 is cooled to 
75° F and determine: 

(a) the final dew point temperature Ans. 72.3° F 

(b) the final wet bulb temperature Ans. 73° F 

(c) the final relative humidity Ans. 92 % 
id) the final total heat per pound of air 

Ans. 36.6 Btu/lb 

10. Assume that the air in Problem 8 is cooled 
to 55° F and determine: 

(a) the total heat removed per pound of air 

Ans. 17.2 Btu/lb 

(6) the sensible heat removed per pound of air 

Ans. 8.40 Btu/lb 

(c) the latent heat removed per pound of air 

Ans. 8.8 Btu/lb 



6 

Refrigeration 
and the Vapor 
Compression 
System 



6-1. Refrigeration. In general, refrigeration 
is denned as any process of heat removal. More 
specifically, refrigeration is denned as that 
branch of science which deals with the process 
of reducing and maintaining the temperature of 
a space or material below the temperature of the 
surroundings. 

To accomplish this, heat must be removed 
from the body being refrigerated and transferred 
to another body whose temperature is below that 
of the refrigerated body. Since the heat removed 
from the refrigerated body is transferred to 
another body, it is evident that refrigerating and 
heating are actually opposite ends of the same 
process. Often only the desired result distin- 
guishes one from the other. 
6-2. Need for Thermal Insulation. Since 
heat will always travel from a region of high 
temperature to a region of lower temperature, 
there is always a continuous flow of heat into 
the refrigerated region from the warmer sur- 
roundings. To limit the flow of heat into the 
refrigerated region to some practical minimum, 
it is usually necessary to isolate the region from 
its surroundings with a good heat insulating 
material. 

6-3. The Heat Load. The rate at which heat 
must be removed from the refrigerated space or 



material in order to produce and maintain the 
desired temperature conditions is called the heat 
load. In most refrigerating applications the total 
heat load on the refrigerating equipment is the 
sum of the heat that leaks into the refrigerated 
space through the insulated walls, the heat that 
enters the space through door openings, and the 
heat that must be removed from the refrigerated 
product in order to reduce the temperature of 
the product to the space or storage conditions. 
Heat given off by people working in the re- 
frigerated space and by motors, lights, and other 
electrical equipment also contributes to the load 
on the refrigerating equipment. 

Methods of calculating the heat load are 
discussed in Chapter 10. 
6-4. The Refrigerating Agent. In any re- 
frigerating process the body employed as the 
heat absorber or cooling agent is called the 
refrigerant. 

All cooling processes may be classified as 
either sensible or latent according to the effect 
the absorbed heat has upon the refrigerant. 
When the absorbed heat causes an increase in 
the temperature of the refrigerant, the cooling 
process is said to be sensible, whereas when the 
absorbed heat causes a change in the physical 
state of the refrigerant (either melting or vapor- 
izing), the cooling process .is said to be latent. 
With either process, if the refrigeratingeffectisto 
be continuous, the temperature of the refriger- 
ating agent must be maintained continuously 
below that of the space or material being 
refrigerated. 

To illustrate, assume that 1 lb of water at 
32° F is placed in an open container inside an 
insulated space having an initial temperature of 
70° F (Fig. 6-1). For a time, heat will flow from 
the 70° F space into the 32° F water and the 
temperature of the space will decrease. How- 
ever, for each one Btu of heat that the water 
absorbs from the space, the temperature of the 
water will increase 1° F, so that as the tempera- 
ture of the space decreases, the temperature of 
the water increases. Soon the temperatures of 
the water and the space will be exactly die same 
and no heat transfer will take place. Refrigera- 
tion will not be continuous because the tempera- 
ture of the refrigerant does not remain below the 
temperature of the space being refrigerated. 

Now assume that 1 lb of ice, also at 32° F, is 
substituted for the water (Fig. 6-2). This time 



71 



72 PRINCIPLES OF REFRIGERATION 



Insulation 




Heat leaking 
through insulation 

Fig. 6-1. Heat flows from warm space to cold water. 
Water temperature rises as space temperature 
decreases. Refrigeration will not be continuous. 



the temperature of the refrigerant does not 
change as it absorbs heat from the space. The 
ice merely changes from the solid to the liquid 
state while its temperature remains constant at 
32° F. The heat absorbed by the ice leaves the 
space in the water going out the drain and the 
refrigerating effect will be continuous until all 
the ice has melted. 

It is both possible and practical to achieve 
continuous refrigeration with a sensible cooling 
process provided that the refrigerant is con- 
tinuously chilled and recirculated through the 
refrigerated space as shown in Fig. 6-3. 

Latent cooling may be accomplished with 
either solid or liquid refrigerants. The solid 
refrigerants most frequently employed are ice 
and solid carbon dioxide (dry ice). Ice, of 
course, melts into the liquid phase at 32° F, 
whereas solid carbon dioxide sublimes directly 
into the vapor phase at a temperature of —109° F 
under standard atmospheric pressure. 
6-5. Ice Refrigeration. Melting ice has been 
used successfully for many years as a refrigerant. 
Not too many years ago ice was the only cooling 
agent available for use in domestic and small 
commercial refrigerators. 

In a typical ice refrigerator (Fig. 6-4) the heat 
entering the refrigerated space from all the 
various sources reaches the melting ice primarily 
by convection currents set up in the air of the 
refrigerated space. The air in contact with the 
warm product and walls of the space is heated 



by heat conducted to it from these materials. 
As the air is warmed it expands and rises to the 
top of the space carrying the heat with it to the 
ice compartment. In passing over the ice the air 
is cooled as heat is conducted from the air to the 
ice. On cooling, the air becomes more dense and 
falls back into the storage space, whereupon it 
absorbs more heat and the cycling continues. 
The air in carrying the heat from the warm walls 
and stored product to the melting ice acts as a 
heat transfer agent. 

To insure adequate air circulation within the 
refrigerated space, the ice should be located near 
the top of the refrigerator and proper baffling 
should be installed to provide direct and un- 
restricted paths of air flow. A drip pan must be 
located beneath the ice to collect the water which 
results from the melting. 

Ice has certain disadvantages which tend to 
limit its usefulness as a refrigerant. For instance, 
with ice it is not possible to obtain the low tem- 
peratures required in many refrigeration applica- 
tions. Ordinarily, 32° F is the minimum tem- 
perature obtainable through the melting of ice 
alone. In some cases, the melting temperature 
of the ice can be lowered to approximately 0° F 
by adding sodium chloride or calcium chloride 
to produce a freezing mixture. 

Some of the other more obvious disadvantages 
of ice are the necessity of frequently replenishing 
the supply, a practice which is neither convenient 
nor economical, and the problem of disposing of 
the water resulting from the melting. 



Insulation 




Heat leaking 
through insulation 

Fig. 6-2. Heat flows from warm space to cold ice. 
Temperature of space decreases as ice melts. Tem- 
perature of ice remains at 32° F. Heat absorbed by 
ice leaves space in water going out the drain. 



REFRIGERATION AND THE VAPOR COMPRESSION SYSTEM 73 



Another less obvious, but more important, 
disadvantage of employing ice as a refrigerant is 
the difficulty experienced in controlling the rate 
of refrigeration, which in turn makes it difficult 
to maintain the desired low temperature level 
within the refrigerated space. Since the rate at 
which the ice absorbs heat is directly propor- 
tional to the surface area of the ice and to the 



6-6. Liquid Refrigerants. The ability of 
liquids to absorb enormous quantities of heat as 
they vaporize is the basis of the modern mechani- 
cal refrigerating system. As refrigerants, vapor- 
izing liquids have a number of advantages over 
melting solids in that the vaporizing process is 
more easily controlled, that is, the refrigerating 
effect can be started and stopped at will, the rate 




Fig. 6-3. Continuous sensible cooling. Heat taken in by the water in the space is given up to the ice. 



temperature difference between the space tem- 
perature and the melting temperature of the ice, 
the rate of heat absorption by the ice diminishes 
as the surface area of the ice is diminished by the 
melting process. Naturally, when the refriger- 
ating rate diminishes to the point that the heat 
is not being removed at the same rate that it is 
accumulating in the space from the various heat 
sources, the temperature of the space will 
increase. 

Despite its disadvantages, ice is preferable to 
mechanical refrigeration in some applications. 
Fresh vegetables, fish, and poultry are often 
packed and shipped in cracked ice to prevent 
dehydration and to preserve appearance. Too, 
ice has tremendous eye appeal and can be used 
to considerable advantage in the displaying and 
serving of certain foods such as salads, cocktails, 
etc., and in chilling beverages. 



of cooling can be predetermined within small 
limits, and the vaporizing temperature of the 
liquid can be governed by controlling the pres- 
sure at which the liquid vaporizes. Moreover, 
the vapor can be readily collected and condensed 
back into the liquid state so that the same liquid 
can be used over and over again to provide a 
continuous supply of liquid for vaporization. 

Until now, in discussing the various properties 
of fluids, water, because of its familiarity, has 
been used in all examples. However, because of 
its relatively high saturation temperature, and 
for other reasons, water is not suitable for use as 
a refrigerant in the vapor-compression cycle. In 
order to vaporize at temperatures low enough to 
satisfy most refrigeration requirements, water 
would have to vaporize under very low pres- 
sures, which are difficult to produce and main- 
tain economically. 



74 PRINCIPLES OF REFRIGERATION 



v;;;;;;;m//m;// ;;;;;;m/m/;;//. 






53£ 



,1 fs^i\ - 

I V 40* Ti / an° » . 



\ 



^ Drain 



44' 



WW///////MWA 









I 

? - Baffle 



Fig. 6-4. Ice refrigerator. Heat is carried from 
warm walls and product to the ice by air circulation 
within the refrigerated space. Air circulation is by 
gravity. 



There are numerous other fluids which have 
lower saturation temperatures than water at the 
same pressure. However, many of these fluids 
have other properties that render them unsuitable 
for use as refrigerants. Actually, only a relatively 
few fluids have properties that make them desir- 
able as refrigerants, and most of these have been 
compounded specially for that purpose. 

There is no one refrigerant which is best suited 
for all the different applications and operating 
conditions. For any specific application the 
refrigerant selected should be the one whose 
properties most closely fit the particular require- 
ments of the application. 

Of all of the fluids now in use as refrigerants, 
the one fluid which most nearly meets all the 
qualifications of the ideal general-purpose re- 
frigerant is a fluorinated hydrocarbon of the 
methane series having the chemical name di- 
chlorodifluoromethane (CC1 2 F 2 ). It is one of a 
group of refrigerants introduced to the industry 
under the trade name of "Freon," but is now 
manufactured under several other proprietary 
designations. To avoid the confusion inherent 
jn the use of proprietory or chemical names, this 
compound is now referred to as Refrigerant-12. 
Refrigerant-12 (R-12) has a saturation tempera- 
ture of — 21. 6° F at standard atmospheric 
pressure. For this reason, R-12 can be stored as 



a liquid at ordinary temperatures only if confined 
under pressure in heavy steel cylinders. 

Table 16-3 is a tabulation of the thermo- 
dynamic properties of R-12 saturated liquid and 
vapor. This table lists, among other things, the 
saturation temperature of R-12 corresponding 
to various pressures. Tables 16-4 through 16-6 
list the thermodynamic properties of some of the 
other more commonly used refrigerants. These 
tables are similar to the saturated liquid and 
vapor tables previously discussed and are em- 
ployed in the same manner. 
6-7. Vaporizing the Refrigerant. An in- 
sulated space can be adequately refrigerated 
by merely allowing liquid R-12 to vaporize in a 
container vented to the outside as shown in Fig. 
6-5. Since the R-12 is under atmospheric pres- 
sure, its saturation temperature is — 21. 6° F. 
Vaporizing at this low temperature, the R-12 
readily absorbs heat from the 40° F space 
through the walls of the containing vessel. The 
heat absorbed by the vaporizing liquid leaves the 
space in the vapor escaping through the open 
vent. Since the temperature of the liquid 
remains constant during the vaporizing process, 
refrigeration will continue until all the liquid is 
vaporized. 

Any container, such as the one in Fig. 6-5, in 



js^v, 



Refrigerant vapor 
/ at atmospheric 
( pressure 




Fig. 6-5. The Refrigerant-12 liquid vaporizes as it 
takes in heat from the 40° F space. The heat taken in 
by the refrigerant leaves the space in the vapor 
escaping through the vent. 



REFRIGERATION AND THE VAPOR COMPRESSION SYSTEM 75 



which a refrigerant is vaporized during a re- 
frigerating process is called an evaporator and 
is one of the essential parts of any mechanical 
refrigerating system. 

6-8. Controlling the Vaporizing Tempera- 
ture. The temperature at which the liquid 
vaporizes in the evaporator can be controlled by 
controlling the pressure of the vapor over the 
liquid, which in turn is governed by regulating 
the rate at which the vapor escapes from the 
evaporator (Section 4-5). For example, if a hand 
valve is installed in the vent line and the vent is 
partially closed off so that the vapor cannot 
escape freely from the evaporator, vapor will 
collect over the liquid causing the pressure in the 
evaporator to rise with a corresponding increase 
in the saturation temperature of the refrigerant 
(Fig. 6-6). By carefully adjusting the vent valve 
to regulate the flow of vapor from the evaporator, 
it is possible to control die pressure of the vapor 
over the liquid and cause the R-12 to vaporize at 
any desired temperature between —21.6° F and 
the space temperature. Should the vent valve be 
completely closed so that no vapor is allowed to 
escape from the evaporator, the pressure in the 
evaporator will increase to a point such that the 
saturation temperature of the liquid will be equal 
to the space temperature, or 40° F. When this 
occurs, there will be no temperature differential 



Refrigerant 

vapor above 

atmospheric 

pressure 



Refrigerant-12 
liquid boiling at 30*F 





Pressure of 

refrigerant vapor 

below atmospheric 



Refrigerant-12 liquid 
'boiling at -100"F 



Fig. 6-6. The boiling temperature of the liquid 
refrigerant in the evaporator is controlled by 
controlling the pressure of the vapor over the liquid 
with the throttling valve in the vent. 



Fig. 6-7. Pressure of refrigerant in evaporator 
reduced below atmospheric by action of a vapor 
pump. 

and no heat will flow from the space to the 
refrigerant. Vaporization will cease and no 
further cooling will take place. 

When vaporizing temperatures below —21.6° 
F are required, it is necessary to reduce the 
pressure in the evaporator to some pressure 
below atmospheric. This can be accomplished 
through the use of a vapor pump as shown in 
Fig. 6-7. By this method, vaporization of the 
liquid R-12 can be brought about at very low 
temperatures in accordance with the pressure- 
temperature relationships given in Table 16-3. 
6-9. Maintaining a Constant Amount of 
Liquid in the Evaporator. Continuous 
vaporization of the liquid in the evaporator 
requires that the supply of liquid be continuously 
replenished if the amount of liquid in the 
evaporator is to be maintained constant. One 
method of replenishing the supply of liquid in 
the evaporator is through the use of a float valve 
assembly as illustrated in Fig. 6-8. The action 
of the float assembly is to maintain a constant 
level of liquid in the evaporator by allowing 
liquid to flow into the evaporator from the 
storage tank or cylinder at exactly the same rate 
that the supply of liquid in the evaporator is 
being depleted by vaporization. Any increase in 
the rate of vaporization causes the liquid level in 
the evaporator to drop slightly, thereby opening 



76 PRINCIPLES OF REFRIGERATION 




High pressure 
S liquid 
refrigerant 



Needle valve 
~ assembly 



_ Low pressure 
liquid refrigerant 




KaSSsiSfc'' 



Fig. 6-8. Float valve assembly maintains constant 
liquid level in evaporator. The pressure of the 
refrigerant is reduced as the refrigerant passes 
through the needle valve. 



the needle valve wider and allowing liquid to 
flow into the evaporator at a higher rate. Like- 
wise, any decrease in the rate of vaporization 
causes the liquid level to rise slightly, thereby 
moving the needle valve in the closing direction 
to reduce the flow of liquid into the evaporator. 
When vaporization ceases entirely, the rising 
liquid level will close the float valve tightly and 
stop the flow of liquid completely. When vapori- 
zation is resumed, the liquid level will fall allow- 
ing the float valve to open and admit liquid to 
the evaporator. 

The liquid refrigerant does not vaporize in the 
storage cylinder and feed line because the pres- 
sure in the cylinder is such that the saturation 
temperature of the refrigerant is equal to the 
temperature of the surroundings (see Section 
4-10). The high pressure existing in the cylinder 
forces the liquid to flow through the feed line 
and the float valve into the lower pressure 
evaporator. In passing through the float valve, 
the high pressure refrigerant undergoes a pres- 
sure drop which reduces its pressure to the 
evaporator pressure, thereby permitting the re- 
frigerant liquid to vaporize in the evaporator at 
the desired low temperature. 

Any device, such as the float valve illustrated 



in Fig. 6-8, used to regulate the flow of liquid 
refrigerant into the evaporator is called a 
refrigerant flow control. The refrigerant flow 
control is an essential part of every mechanical 
refrigerating system. 

There are five different types of refrigerant 
flow controls, all of which are in use to some 
extent at the present time. Each of these 
distinct types is discussed at length in Chapter 
17. The float type of control illustrated in Fig. 
6-8 has some disadvantages, mainly bulkiness, 
which tend to limit its use to some few special 
applications. The most widely used type of 
refrigerant flow control is the thermostatic 
expansion valve. A flow diagram illustrating the 
use of a thermostatic expansion valve to control 
the flow of refrigerant into a serpentine coil 
type evaporator is shown in Fig. 6-9. 
6-10. Salvaging the Refrigerant. As a matter 
of convenience and economy it is not practical 
to permit the refrigerant vapor to escape to the 
outside and be lost by diffusion into the air. 
The vapor must be collected continuously and 
condensed back into the liquid state so that the 
same refrigerant is used over and over again, 
thereby eliminating the need for ever replenish- 
ing the supply of refrigerant in the system. To 

High pressure 
liquid 




Low pressure 

Hiquid-vapor 

mixture 



Fig. 6-9. Serpentine coil evaporator with thermo- 
static expansion valve refrigerant control. 



REFRIGERATION AND THE VAPOR COMPRESSION SYSTEM 77 



provide some means of condensing the vapor, 
another piece of equipment, a condenser, must 
be added to the system (Fig. 6-10). 

Since the refrigerant vaporizes in the evapor- 
ator because it absorbs the necessary latent heat 
from the refrigerated space, all that is required 
in order to condense the vapor back into the 
liquid state is that the latent heat be caused to 
flow out of the vapor into another body. The 
body of material employed to absorb the latent 
heat from the vapor, thereby causing the vapor 
to condense, is called the condensing medium. 
The most common condensing media are air 
and water. The water used as a condensing 
medium is usually supplied from the city main 
or from a cooling tower. The air used as a 
condensing medium is ordinary outdoor air at 
normal temperatures. 

For heat to flow out of the refrigerant vapor 
into the condensing medium the temperature 
of the condensing medium must be below that 
of the refrigerant vapor. However, since the 
pressure and temperature of the saturated 
vapor leaving the evaporator are the same as 
those of the vaporizing liquid, the temperature 
of the vapor will always be considerably below 
that of any normally available condensing 
medium. Therefore, heat will not flow out of 
the refrigerant vapor into the air or water used 
as the condensing medium until the saturation 
temperature of the refrigerant vapor has been 
increased by compression to some temperature 
above the temperature of the condensing 
medium. The vapor pump or compressor 
shown in Fig. 6-10 serves this purpose. 

Before compression, the refrigerant vapor is 
at the vaporizing temperature and pressure. 
Since the pressure of the vapor is low, the 
corresponding saturation temperature is also 
low. During compression the pressure of the 
vapor is increased to a point such that the 
corresponding saturation temperature is above 
the temperature of the condensing medium being 
employed. At the same time, since mechani- 
cal work is done on the vapor in compressing 
it to the higher pressure, the internal energy of 
the vapor is increased with a corresponding 
increase in the temperature of the vapor. 

After compression, the high-pressure, high- 
temperature vapor is discharged into the con- 
denser where it gives up heat to the lower 
temperature condensing medium. Since a vapor 



L 



low-pressure, 
low-temperature 



Low-pressure, 

low-temperature 

liquid-vapor mixture 

Refrigerant control -s / 



high-temperature 




High-pressure, 
high-tempenhii 



J 



High-pressure, . 
high-temperature' 
liquid-vapor mixture 



High-pressure, 

Mgh-temperitun! - 

liquid 



Fig. 6-10. Collecting and condensing the refrigerant 
vapor. Refrigerant absorbs heat in evaporator and 
gives off heat in the condenser. 

cannot be cooled to a temperature below its 
saturation temperature, the continuous loss of 
heat by the refrigerant vapor in the condenser 
causes the vapor to condense into the liquid 
state at the new, higher pressure and saturation 
temperature. The heat given off by the vapor in 
the condenser is carried away by the condensing 
medium. The resulting condensed liquid, whose 
temperature and pressure will be the same as 
those of the condensing vapor, flows out of the 
condenser into the liquid storage tank and is 
then ready to be recirculated to the evaporator. 
Notice that the refrigerant, sometimes called 
the working fluid, is merely a heat transfer 
agent which carries the heat from the refriger- 
ated space to the outside. The refrigerant 
absorbs heat from the refrigerated space in the 
evaporator, carries it out of the space, and 
rejects it to the condensing medium in the 
condenser. 



78 PRINCIPLES OF REFRIGERATION 



6-11. Typical Vapor-Compression System. 

A flow diagram of a simple vapor-compression 
system is shown in Fig. 6-11. The principal 
parts of the system are: (1) an evaporator, 
whose function it is to provide a heat transfer 
surface through which heat can pass from the 
refrigerated space or product into the vaporizing 
refrigerant; (2) a suction line, which conveys 
the low pressure vapor from the evaporator 
to the suction inlet of the compressor; (3) a 
vapor compressor, whose function it is to 
remove the vapor from the evaporator, and to 
raise the temperature and pressure of the vapor 
to a point such that the vapor can be condensed 
with normally available condensing media; (4) 
a "hot-gas" or discharge line which delivers the 
high-pressure, high-temperature vapor from the 
discharge of the compressor to the condenser; 
(5) a condenser, whose purpose it is to provide 
a heat transfer surface through which heat 
passes from the hot refrigerant vapor to the 
condensing medium; (6) a receiver tank, which 
provides storage for the liquid condenser so that 
a constant supply of liquid is available to the 
evaporator as needed; (7) a liquid line, which 
carries the liquid refrigerant from the receiver 
tank to the refrigerant flow control; (8) a 
refrigerant flow control, whose function it is to 
meter the proper amount of refrigerant to the 
evaporator and to reduce the pressure of the 
liquid entering the evaporator so that the liquid 
will vaporize in the evaporator at the desired low 
temperature. 



4-12. Service Valves. The suction and dis- 
charge sides of the compressor and the outlet 
of the receiver tank are usually equipped with 
manual shut-off valves for use during service 
operations. These valves are known as the 
"suction service valve," the discharge service 
valve," and the "receiver tank valve," respec- 
tively. Receiver tanks on large systems frequent- 
ly have shut-off valves on both the inlet and the 
outlet. 

6-13. Division of the System. A refriger- 
ating system is divided into two parts according 
to the pressure exerted by the refrigerant in the 
two parts. The low pressure part of the system 
consists of the refrigerant flow control, the 
evaporator, and the suction line. The pressure 
exerted by the refrigerant in these parts is the 
low pressure under which the refrigerant is 
vaporizing in the evaporator. This pressure is 
known variously as the "low side pressure," 
the "evaporator pressure," the "suction pres- 
sure," or the "back pressure." During service 
operations this pressure is usually measured 
at the compressor by installing a compound 
gage on the gage port of the suction service 
valve. 

The high pressure side or "high side" of the 
system consists of the compressor, the discharge 
or "hot gas" line, the condenser, the receiver 
tank, and the liquid line. The pressure exerted 
by the refrigerant in this part of the system is 
the high pressure under which the refrigerant is 
condensing in the condenser. This pressure is 



Suction 
line 



ine~V. 
® 



Refrigerant 
flow control"! 



Evaporator-] 



Suction _ 
service valve 

Compressor 




■<8K 



j 



"SB", rDischargeline 

valve | (g) (-Condenser 



Liquid 

dj~ ,ine 



Receiver 
J/tank valve 



® 
VJteceiver 
tank 



Fig. 6-11. Flow diagram of simple 
vapor compression system show- 
ing the principal parts. 



REFRIGERATION AND THE VAPOR COMPRESSION SYSTEM 79 



Fig, 6-12. Air-cooled condens- 
ing unit. Note fan mounted on 
motor shaft to circulate air 
over condenser. 



Receiver 
tank 



Compressor 




ir- coo led 
condenser 



Compressor 
driver 



called the "condensing pressure," the "dis- 
charge pressure," or, more often, the "head 
pressure." 

The dividing points between the high and low 
pressure sides of the system are the refrigerant 
flow control, where the pressure of the refriger- 
ant is reduced from the condensing pressure to 
the vaporizing pressure, and the discharge valves 
in the compressor, through which the high 
pressure vapor is exhausted after compression.* 
It should be noted that, although the compressor 
is considered to be a part of the high side of the 
system, the pressure on the suction side of the 
compressor and in the crankcase is the low side 
pressure. The change in pressure, of course, 
occurs in the cylinder during the compression 
process. 

6-14, Condensing Units. The compressor, 
hot gas line, condenser, and receiver tank, 
along with the compressor driver (usually an 
electric motor), are often combined into one 
compact unit as shown in Fig. 6-12. Such an 
assembly is called a condensing unit because its 
function in the system is to reclaim the vapor 
and condense it back into the liquid state. 

Condensing units are often classified accord- 

• Care should be taken not to confuse the suction 
and discharge valves in the compressor with the 
suction and discharge service valves. The suction 
and discharge valves in a reciprocating compressor 
perform the same function as the intake and exhaust 
valves in an automobile engine and are vital to the 
operation of the compressor, whereas the suction and 
discharge service valves serve no useful purpose inso- 
far as the operation of the compressor is concerned. 
The latter valves are used only to facilitate service 
operations, as their nomenclature implies. 



ing to condensing medium used to condense the 
refrigerant. A condensing unit employing air as 
the condensing medium (Fig. 6-12) is called an 
air-cooled condensing unit, whereas one employ- 
ing water as the condensing medium is a water- 
cooled condensing unit. 

6-IS. Hermetic Motor-Compressor Assem- 
blies. Condensing units of small horsepower 
are often equipped with hermetically sealed 
motor-compressor assemblies. The assembly 
consists of a direct-driven compressor mounted 
on a common shaft with the motor rotor and 
the whole assembly hermetically sealed in a 
welded steel shell (Fig. 6-13). 

Condensing units equipped with hermetically 
sealed motor-compressor assemblies are known 
as "hermetic condensing units" and are em- 
ployed on a number of small commercial 
refrigerators and on almost all household 
refrigerators, home freezers, and window air 
conditioners. For reasons that will be shown 
later, many hermetic condensing units are not 
equipped with receiver tanks, 

A variation of the hermetic motor-compressor 
assembly is the "accessible hermetic." It is 
similar to the full hermetic except that the shell 
enclosing the assembly is bolted together rather 
than seam welded. (Fig. 6-14). The bolted 
construction permits the assemblies to be 
opened in the field for servicing. 
6-16. Definition of a Cycle. As the refriger- 
ant circulates through the system, it passes 
through a number of changes in state or 
condition, each of which is called a process. 
The refrigerant starts at some initial state or 
condition, passes through a series of processes 
in a definite sequence, and returns to the initial 



80 PRINCIPLES OF REFRIGERATION 




Fig. 4-13. Air-cooled condensing unit employing hermetic motor-compressor. Note separate fan to circulate 
air over condenser. (Courtesy Tecumseh Products Company.) 



condition. This series of processes is called a 
cycle. The simple vapor-compression refriger- 
ation cycle is made up of four fundamental 
processes: (1) expansion, (2) vaporization, (3) 
compression, and (4) condensation. 

To understand properly the refrigeration 
cycle it is necessary to consider each process in 
the cycle both separately and in relation to the 
complete cycle. Any change in any one process 
in the cycle will bring about changes in all the 
other processes in the cycle. 
6-17. Typical Vapor-Compress ton Cycle. 
A typical vapor-compression cycle is shown in 
Fig. 6-15. Starting at the receiver tank, high- 
temperature, high-pressure liquid refrigerant 
flows from the receiver tank through the liquid 
line to the refrigerant flow control. The 
pressure of the liquid is reduced to the evapor- 
ator pressure as the liquid passes through the 
refrigerant flow control so that the saturation 
temperature of the refrigerant entering the 
evaporator will be below the temperature of the 
refrigerated space. It will be shown later that a 



part of the liquid vaporizes as it passes through 
the refrigerant control in order to reduce the 
temperature of the liquid to the evaporating 
temperature. 

In the evaporator, the liquid vaporizes at a 
constant pressure and temperature as heat to 
supply the latent heat of vaporization passes 
from the refrigerated space through the walls of 
the evaporator to the vaporizing liquid. By the 
action of the compressor, the vapor resulting 
from the vaporization is drawn from the 
evaporator through the suction line into the 
suction inlet of the compressor. The vapor 
leaving the evaporator is saturated and its 
temperature and pressure are the same as those 
of the vaporizing liquid. While flowing through 
the suction line from the evaporator to the 
compressor, the vapor usually absorbs heat 
from the air surrounding the suction line and 
becomes superheated. Although the tempera- 
ture of the vapor increases somewhat in the 
suction line as the result of superheating, the 
pressure of the vapor does not change so that 



REFRIGERATION AND THE VAPOR COMPRESSION SYSTEM 81 




o 

u 

I 

8 



r 

I 

5 

8 

E 

II 
!i 

« 
£ 

L- 
C V 

I* 

9- 3 
£f 

2<5 

■i • 



82 PRINCIPLES OF REFRIGERATION 



Liquid-vapor mixture 
30'F-28.46 psig 



Liquid-vapor mixture 
30T-28.46 psig " 

Saturated vapor 
30°F-28.46 psig 

Superheated vapor 
132*^-120.6 psig" 

Superheated vapo r t ^ >-> 

70°F-28.46 psig~Xi£l(£ 




Subcooled liquid 
86°F-120.6 psig 



Saturated vapor 
102"F-120.6 psig 



Liquid-vapor mixture 
" 102°F-120.6 psig 



Saturated liquid 

102°F-120.6 psig 

Fig. 6-15. Typical refrigeration system showing the condition of the refrigerant at various points. 



the pressure of the vapor entering the compres- 
sor is the same as the vaporizing pressure.* 

In the compressor, the temperature and 
pressure of the vapor are raised by compression 
and the high-temperature, high-pressure vapor 
is discharged from the compressor into the 
hot-gas line. The vapor flows through the 
hot-gas line to the condenser where it gives up 
heat to the relatively cool air being drawn 
across the condenser by the condenser fan. 
As the hot vapor gives off heat to the cooler air, 
its temperature is reduced to the new saturation 
temperature corresponding to its new pressure 
and the vapor condenses back into the liquid 
state as additional heat is removed. By the 
time the refrigerant reaches the bottom of the 
condenser, all of the vapor is condensed and the 
liquid passes into the receiver tank, ready to be 
recirculated. 

6-18. The Compression Process. In 
modern, high speed compressors, compression 
takes place very rapidly and the vapor is in 
contact with the compressor cylinder for only a 
short time. Because the time of compression 
is short and because the mean temperature 

* Actually, the pressure of the vapor decreases 
slightly between the evaporator and compressor 
because of the friction loss in the suction line 
resulting from the flow. 



differential between the refrigerant vapor and 
the cylinder wall is small, the flow of heat 
either to or from the refrigerant during com- 
pression is usually negligible. Therefore, com- 
pression of the vapor in a refrigeration com- 
pressor is assumed to occur adiabatically. 

Although no heat as such is transferred either 
to or from the refrigerant during the compres- 
sion, the temperature and enthalpy of the vapor 
are increased because of the mechanical work 
done on the vapor by the piston. Whenever a 
vapor is compressed, unless the vapor is cooled 
during the compression, the internal kinetic 
energy of the vapor is increased by an amount 
equal to the amount of work done on the vapor 
(Section 3-12). Therefore, when a vapor is 
compressed adiabatically, as in a refrigeration 
compressor, wherein no heat is removed from 
the vapor during the compression, the tempera- 
ture and enthalpy are increased in direct 
proportion to the amount of work done during 
the compression. The greater the work of 
compression, the greater is the increase in 
temperature and enthalpy. 

The energy equivalent of the work done is 
called the heat of compression. The energy to 
do the work of compression, which is trans- 
ferred to the vapor during the compression 
process, is supplied by the compressor driver, 



REFRIGERATION AND THE VAPOR COMPRESSION SYSTEM 83 



usually an electric motor. It will be shown 
later that the theoretical horsepower required 
to drive the compressor can be calculated from 
the heat of compression. 
6-19. Discharge Temperature. Care should 
be taken not to confuse discharge temperature 
with condensing temperature. The discharge 
temperature is that at which the vapor is dis- 
charged from the compressor, whereas the 
condensing temperature is that at which the 
vapor condenses in the condenser and is the 
saturation temperature of the vapor correspond- 
ing to the pressure in the condenser. Because 
the vapor is usually superheated as it enters the 
compressor and because it contains the heat of 
compression, the vapor discharged from the 
compressor is highly superheated and its 
temperature is considerably above the satura- 
tion temperature corresponding to its pressure. 
The discharge vapor is cooled to the condensing 
temperature as it flows through the hot-gas line 
and through the upper part of the condenser, 
whereupon the further removal of heat from 
the vapor causes the vapor to condense at the 
saturation temperature corresponding to the 
pressure in the condenser. 
6-20. Condensing Temperature. To provide 
a continuous refrigerating effect the refrigerant 
vapor must be condensed in the condenser at 
the same rate that the refrigerant liquid is 
vaporized in the evaporator. This means that 
heat must leave the system at the condenser at 
the same rate that heat is taken into the system 
in the evaporator and suction line, and in the 
compressor as a result of the work of compres- 
sion. Obviously, any increase in the rate of 
vaporization will increase the required rate of 
heat transfer at the condenser. 

The rate at which heat will flow through the 
walls of the condenser from the refrigerant 
vapor to the condensing medium is the function 
of three factors: (1) the area of the condensing 
surface, (2) the coefficient of conductance of the 
condenser walls, and (3) the temperature differ- 
ence between the refrigerant vapor and the 
condensing medium. For any given condenser, 
the area of the condensing surface and the 
coefficient of conductance are fixed so that the 
rate of heat transfer through the condenser 
walls depends only on the temperature differ- 
ence between the refrigerant vapor and the 
condensing medium. 



Since the condensing temperature is always 
equal to the temperature of the condensing 
medium plus the temperature difference between 
the condensing refrigerant and the condensing 
medium, it follows that the condensing tempera- 
ture varies directly with the temperature of the 
condensing medium and with the required rate 
of heat transfer at the condenser. 
6-21. Condensing Pressure. The condensing 
pressure is always the saturation pressure 
corresponding to the temperature of the liquid- 
vapor mixture in the condenser. 

When the compressor is not running, the 
temperature of the refrigerant mixture in the 
condenser will be the same as that of the 
surrounding air, and the corresponding satura- 
tion pressure will be relatively low. Conse- 
quently, when the compressor is started, the 
vapor pumped over into the condenser will not 
begin to condense immediately because there is 
no temperature differential between the refriger- 
ant and the condensing medium, and therefore 
no heat transfer between the two. Because of 
the throttling action of the refrigerant control, 
the condenser may be visualized as a closed 
container, and as more and more vapor is 
pumped into the condenser without condensing, 
the pressure in the condenser increases to a 
point where the saturation temperature of 
vapor is sufficiently high to permit the required 
rate of heat transfer between the refrigerant 
and the condensing medium. When the required 
rate of heat transfer is reached, the vapor will 
condense as fast as it is pumped into the 
condenser, whereupon the pressure in the con- 
denser will stabilize and remain more or less 
constant during the balance of the running cycle. 
6-22. Refrigerating Effect. The quantity of 
heat that each pound of refrigerant absorbs 
from the refrigerated space is known as the 
refrigerating effect. For example, when 1 lb of 
ice melts it will absorb from the surrounding air 
and from adjacent objects an amount of heat 
equal to its latent heat of fusion. If the ice melts 
at 32° F it will absorb 144 Btu per pound, so that 
the refrigerating effect of 1 lb of ice is 144 Btu. 

Likewise, when a liquid refrigerant vaporizes 
as it flows through the evaporator it will absorb 
an amount of heat equal to that required to 
vaporize it; thus the refrigerating effect of 1 lb 
of liquid refrigerant is potentially equal to its 
latent heat of vaporization. If the temperature 



84 PRINCIPLES OF REFRIGERATION 



of the liquid entering the refrigerant control 
from the liquid line is exactly equal to the 
vaporizing temperature in the evaporator the 
entire pound of liquid will vaporize in the 
evaporator and produce useful cooling, in 
which case the refrigerating effect per pound 
of refrigerant circulated will be equal to the 
total latent heat of vaporization. However, in 
an actual cycle the temperature of the liquid 
entering the refrigerant control is always 
considerably higher than the vaporizing tem- 
perature in the evaporator, and must first be 
reduced to the evaporator temperature before 
the liquid can vaporize in the evaporator and 
absorb heat from the refrigerated space. For 
this reason, only a part of each pound of Hquid 
actually vaporizes in the evaporator and 
produces useful cooling. Therefore, the refriger- 
ating effect per pound of liquid circulated is 
always less than the total latent heat of vapori- 
zation. 

With reference to Fig. 6-15, the pressure of 
the vapor condensing in the condenser is 120 
psig and the condensing temperature (satur- 
ation temperature) of the R-12 vapor corre- 
sponding to this pressure is 102° F. Since 
condensation occurs at a constant temperature, 
the temperature of the liquid resulting from the 
condensation is also 102° F. After condensation, 
as the liquid flows through the lower part of 
the condenser it continues to give up heat to the 
cooler condensing medium, so that before the 
liquid leaves the condenser its temperature is 
usually reduced somewhat below the tempera- 
ture at which it condensed. The liquid is then 
said to be subcooled. The temperature at 
which the liquid leaves the condenser depends 
upon the temperature of the condensing 
medium and upon how long the hquid remains 
in contact with the condensing medium after 
condensation. 

The liquid may be further subcooled in the 
receiver tank and in the liquid line by surrender- 
ing heat to the surrounding air. In any case, 
because of the heat exchange between the 
refrigerant in the liquid line and the surrounding 
air, the temperature of the liquid approaching 
the refrigerant control is likely to be fairly 
close to the temperature of the air surrounding 
the liquid line. In Fig. 6-1 5, the liquid ap- 
proaches the refrigerant control at a temperature 
of 86° F, whereas its pressure is still the same as 



the condensing pressure, 120.6 psig. Since the 
saturation temperature corresponding to 120.6 
psig is 102° F, the 86° F liquid at the refrigerant 
control is subcooled 16° F (102 - 86) below its 
saturation temperature. 

Since the saturation pressure corresponding 
to 86° F is 93.2 psig, the R-12 can exist in the 
liquid state as long as its pressure is not reduced 
below 93.2 psig. However, as the liquid passes 
through the refrigerant control its pressure is 
reduced from 120.6 psig to 28.46 psig, the 
saturation pressure corresponding to the 30° 
vaporizing temperature of the refrigerant in the 
evaporator. Since the R-12 cannot exist as a 
liquid at any temperature above the saturation 
temperature of 30° F when its pressure is 28.46 
psig, the hquid must surrender enough heat 
to cool itself from 86° F to 30° F at the instant 
that its pressure is reduced in passing through 
the refrigerant control. 

From Table 16-3, the enthalpy of hquid at 
86° F and at 30° F is 27.73 Btu per pound and 
14.76 Btu per pound, respectively, so that each 
pound of liquid must surrender 12.97 Btu 
(27.73 - 14,76) in order to cool from 86° F to 
30° F. Because the hquid expands through the 
refrigerant control so rapidly, the liquid is not 
in contact with the control for a sufficient length 
of time to permit this amount of heat to be 
transferred from the refrigerant to the control. 
Therefore, a portion of each pound of hquid 
vaporizes as the liquid passes through the con- 
trol, and the heat to supply the latent heat of 
vaporization for the portion that vaporizes is 
drawn from the body of the hquid, thereby 
reducing the temperature of the refrigerant to the 
evaporator temperature. In this instance, enough 
of each pound of liquid vaporizes while passing 
through the refrigerant control to absorb exactly 
the 12.97 Btu of sensible heat that each pound of 
hquid must surrender in order to cool from 86° F 
to 30° F and the refrigerant is discharged from 
the refrigerant control into the evaporator as a 
liquid-vapor mixture. Obviously, only the liquid 
portion of the liquid-vapor mixture will vaporize 
in the evaporator and produce useful cooling. 
That portion of each pound of hquid circulated 
which vaporizes in the refrigerant control pro- 
duces no useful cooling and represents a loss of 
refrigerating effect. It follows, then, that the 
refrigerating effect per pound of hquid circulated 
is equal to the total latent heat of vaporization 



REFRIGERATION AND THE VAPOR COMPRESSION SYSTEM 85 



less the amount of heat absorbed by that part of 
each pound that vaporizes in the control to 
reduce the temperature of the liquid to the 
vaporizing temperature. 

From Table 16-3, the latent heat of vaporiza- 
tion of 'R-12 at 30° F is 66.85 Btu per pound. 
Since the loss of refrigerating effect is 12.97 Btu 
per pound, the refrigerating effect in this instance 
is (66.85 - 12.97) 53.88 Btu per pound. 

The percentage of each pound of refrigerant 
that vaporizes in the refrigerant control can be 
determined by dividing the total latent heat of 
vaporization into the heat absorbed by that part 
of the pound that vaporizes in the control. In 
this instance the percentage of each pound 
vaporizing in the control is (12.97/66.85 x 100) 
19.4%. Only 80.6% of each pound circulated 
actually vaporizes in the evaporator and pro- 
duces useful cooling (66.85 x 0.806 = 53.88 
Btu/lb). 

Even though a portion of each pound cir- 
culated vaporizes as it passes through the re- 
frigerant control, the enthalpy of the refrigerant 
does not change in the control. That is, since 
there is no heat transfer between the refrigerant 
and the control, the enthalpy of the liquid-vapor 
mixture discharged from the control into the 
evaporator is exactly the same as the enthalpy of 
the liquid approaching the control. Therefore, 
the difference between the enthalpy of the re- 
frigerant vapor leaving the evaporator and the 
enthalpy of the liquid approaching the control is 
only the amount of heat absorbed by the re- 
frigerant in the evaporator, which is, of course, 
the refrigerating effect. Hence, for any given 
conditions the refrigerating effect per pound can 
be easily determined by subtracting the enthalpy 
of the liquid refrigerant entering the control 
from the enthalpy of the saturated vapor leaving 
the evaporator. 

Example 6-1. Determine the refrigerating 
effect per pound if the temperature of the liquid 
R-12 approaching the refrigerant control is 
86° F and the temperature of the saturated 
vapor leaving the evaporator is 30° F. 

Solution. From Table 
16-3, enthalpy of R-12 satur- 
ated vapor at 30° F = 81.61 Btu/lb 

Enthalpy of R-12 liquid 
at 86° F = 27.73 Btu/lb 

Refrigerating effect per 
pound = 53.88 Btu/lb 



Example 6-2. If, in Example 6-1, the tem- 
perature of the liquid entering the refrigerant 
control is 60° F rather than 86° F, determine the 
refrigerating effect. 

Solution. From Table 
16-3, enthalpy of R-12 
saturated vapor at 30° F =81.61 Btu/lb 

Enthalpy of R-12 liquid 
at 60° F = 21.57 Btu/lb 

Refrigerating effect = 60.04 Btu/lb 

Example 6-3. If, in Example 6-1, the 
pressure in the evaporator is 21.05 psig, and the 
liquid reaching the refrigerant control is 86°" F, 
what is the refrigerating effect? 

Solution. From Table 
16-3, the saturation tem- 
perature of R-12 corre- 
sponding to 21.05 psig is 
20° F and the enthalpy of 
R-12 saturated vapor at that 
temperature = 80.49 Btu/lb 

Enthalpy of R-12 liquid 
at 86° F = 27.73 Btu/lb 

Refrigerating effect = 52.77 Btu/lb 

A comparison of Examples 6-1 and 6-2 in- 
dicates that the refrigerating effect increases as 
the temperature of the liquid approaching the 
refrigerant control decreases, whereas a com- 
parison of Example 6-1 and 6-3 shows that the 
refrigerating effect decreases as the vaporizing 
temperature decreases. Therefore, it is evident 
that the refrigerating effect per pound of liquid 
circulated depends upon two factors: (1) the 
evaporating temperature and (2) the temperature 
at which the liquid refrigerant enters the re- 
frigerant control. The higher the evaporating 
temperature and the lower the temperature of 
the liquid entering the refrigerant control, 
the greater will be the refrigerating effect. 
6-23. System Capacity. The capacity of any 
refrigerating system is the rate at which it will 
remove heat from the refrigerated space and is 
usually stated in Btu per hour or in terms of its 
ice-melting equivalent. 

Before the era of mechanical refrigeration, ice 
was widely used as a cooling medium. With the 
development of mechanical refrigeration, it was 
only natural that the cooling capacity of 
mechanical refrigerators should be compared 
with an ice-melting equivalent. 

When one ton of ice melts.it will absorb 
288,000 Btu (2000 lb x 144 Btu/lb). If one ton 



86 PRINCIPLES OF REFRIGERATION 



of ice melts in one day (24 hr), it will absorb heat 
at the rate of 12,000 Btu/hr (288,000 Btu/24 hr) 
or 200 Btu/min (12,000 Btu/hr/60). Therefore, 
a mechanical refrigerating system having the 
capacity of absorbing heat from the refrigerated 
space at the rate of 200 Btu/min (12,000 Btu/hr) 
is cooling at a rate equivalent to the melting of 
one ton of ice in 24 hr and is said to have a 
capacity of one ton. 

The capacity of a mechanical refrigeration 
system, that is, the rate at which the system will 
remove heat from the refrigerated space, depends 
upon two factors: (1) the weight of refrigerant 
circulated per unit of time and (2) the refrigerat- 
ing effect of each pound circulated. 

Example 6-4. A mechanical refrigerating 
system is operating under conditions such that 
the vaporizing temperature is 30° F while the 
temperature of the liquid approaching the re- 
frigerant control is 86° F. If R-12 is circulated 
through the system at the rate of 5 lb/min, 
determine: 

(«) the refrigerating capacity of the system in 
Btu per hour. 

(b) the refrigerating capacity of the system in 
tons. 



Solution 

(a) From Example 6-1, 

refrigerating effect 
Weight of refrigerant 

circulated per minute 
Refrigerating capacity 

in Btu per minute 

Refrigerating capacity 
in Btu per hour 

(b) Refrigerating capacity 

in tons 



= 53.88 Btu/lb 

= 5 lb 

= 5 x 53.88 

= 269.40 Btu/min 

= 269.40 x 60 
= 16,164 Btu/hr 

_ 269.40 

200 
= 1.347 tons 



6-24. Weight of Refrigerant Circulated per 
Minute per Ton. The weight of refrigerant 
which must be circulated per minute per ton of 
refrigerating capacity for any given operating 
conditions is found by dividing the refrigerating 
effect per pound at the given conditions into 200. 

Example 6-5. An R-12 system is operating 
at conditions such that the vaporizing tempera- 
ture is 20° F and the condensing temperature is 
100° F. If it is assumed that no subcooling of 



the liquid occurs so that the temperature of the 
liquid at the refrigerant control is also 100° F, 
find: 

(a) The refrigerating effect per pound 

(b) The weight of refrigerant circulated per 
minute per ton 

(c) The weight of refrigerant circulated per 
minute for a 10-ton system. 

Solution 

(a) From Table 16-3, en- 

thalpy of R-12 satur- 
ated vapor at 20° F = 80.49 Btu/lb 
Enthalpy of R-12 

liquid at 100° F = 31.16 Btu/lb 

Refrigerating effect 49.33 Btu/lb 

(b) Weight of refrigerant cir- 

culated per minute per 

ton 200 



(c) Weight of refrigerant cir- 
culated per minute for 
a 10-ton system 



49.33 
4.05 lb 



10 x 4.05 
40.5 lb 



Example 6-6. If, in Example 6-5, the liquid 
is subcooled from 100° F to 80° F before it 
reaches the refrigerant control, calculate: 

(a) the refrigerating effect 

(b) the weight of refrigerant circulated per 
minute per ton 

Solution 

(a) From Table 16-3, 

enthalpy of R-12 

saturated vapor at 

20° F 
Enthalpy of R-12 liquid 

at 80° F 
Refrigerating effect 

(b) Weight of refrigerant cir- 

culated per minute per 
ton 



= 80.49 Btu/lb 

= 26J8 Btu/lb 
= 54.21 Btu/lb 
200 



54.21 
= 3.69 lb 



In comparing Examples 6-5 and 6-6, it is 
apparent that the weight of refrigerant which 
must be circulated per minute per ton of re- 
frigerating capacity varies with the refrigerating 
effect and depends upon the operating conditions 
of the system. As the refrigerating effect per 
pound increases, the weight of refrigerant cir- 
culated per minute per ton decreases. 
6-25. Volume of Vapor Displaced per Min- 
ute per Ton. When 1 lb of liquid refrigerant 
vaporizes, the volume of vapor which results 



REFRIGERATION AND THE VAPOR COMPRESSION SYSTEM 87 



depends upon the vaporizing temperature. The 
lower the vaporizing temperature and pressure, 
the greater is the volume of the vapor which is 
produced. When the vaporizing temperature is 
known, the specific volume of the saturated 
vapor which results from the vaporization can 
be found in the saturated vapor tables. For 
instance, from Table 16-3, the specific volume of 
R-12 saturated vapor at 10° F is 1.351 cu ft per 
pound. This means that each pound of R-12 
that vaporizes at 10° F produces 1.351 cuft of 
vapor. Therefore, if 10 lb of R-12 are vaporized 
at 10° F in an evaporator each minute, saturated 
vapor will be produced at the rate of 13.51 cu ft 
per minute (10 x 1.351). 

In order to produce one ton of refrigerating 
capacity, a definite weight of refrigerant must be 
vaporized each minute. The volume of vapor 
which must be removed from the evaporator 
each minute can be calculated by multiplying 
the weight of refrigerant circulated per minute 
by the specific volume of the saturated vapor at 
the vaporizing temperature. 

Example 6-7. Determine the volume of 
vapor to be removed from the evaporator per 
minute per ton of refrigerating capacity for the 
system described in Example 6-5. 

Solution. From 
Table 16-3, specific 
volume of R-12 satur- 
ated vapor at 20° F 

From Example 6-5, 
weight of refrigerant 
circulated per minute 
per ton 

Volume of vapor 
displaced per minute 
per ton 



= 1.121 cu ft/lb 



= 4.05 lb/min/ton 

= 4.05 x 1.121 

= 4.55 cu ft/min/ton 



6-26. Compressor Capacity. In any mecha- 
nical refrigerating system the capacity of the 
compressor must be such that vapor is drawn 
from the evaporator at the same rate that vapor 
is produced by the boiling action of the liquid 
refrigerant. If the refrigerant vaporizes faster 
than the compressor is able to remove the vapor, 
the excess vapor will accumulate in the evapora- 
tor and cause the pressure in the evaporator to 
increase, which in turn will result in raising the 
boiling temperature of the liquid. On the other 
hand, if the capacity of the compressor is such 
that the compressor removes the vapor from the 



evaporator too rapidly, the pressure in the 
evaporator will decrease and result in a decrease 
in the boiling temperature of the liquid. In either 
case, design conditions will not be maintained 
and the refrigerating system will be unsatis- 
factory. 

The maintenance of design conditions and 
therefore good refrigeration depends upon the 
selection of a compressor whose capacity is such 
that the compressor will displace in any given in- 
terval of time a volume of vapor that is equal to 
the volume occupied by the weight of refrig- 
erant which must be vaporized during the same 
time interval in order to produce the required 
refrigerating capacity at the design conditions. 

For instance, in Example 6-7, 4.05 lb of R-12 
must be vaporized each minute at 20° F for each 
one ton of refrigerating capacity desired. In 
vaporizing, the 4.05 lb of R-12 produce 4.55 
cuft of vapor (4.05 x 1.121). If the evaporator 
pressure and the boiling temperature of the 
liquid in the evaporator are to remain constant, 
this volume of vapor must be removed from the 
evaporator each minute for each one ton of 
refrigerating capacity. Hence, the compressor 
selected for a system operating at the conditions 
of Example 6-7 should have a capacity such that 
it will remove vapor from the evaporator at the 
rate of 4.55 cu ft per minute for each ton of 
refrigerating capacity required. For a 10 ton 
system, the compressor would have to remove 
vapor from the evaporator at the rate of 45.50 
cu ft per minute (10 x 4.55). 

PROBLEMS 

1. The temperature of liquid R-12 entering the 
refrigerant control is 86° F and the vaporizing 
temperature 30° F. Determine: 

(a) The refrigerating effect per pound of 
refrigerant circulated. Ans. 53.89 Btu/lb 

(6) The loss of refrigerating effect per pound. 

Ans. 12.96 Btu/lb 

(c) The weight of refrigerant circulated per 
minute per ton. Ans. 3.71 lb/min/ton 

(d)The volume of vapor displaced per 
minute per ton. Ans. 3.48 cu ft/min/ton 

2. If saturated R-12 liquid reaches the refriger- 
ant control at a pressure of 136 psig and the 
vaporizing pressure in the evaporator is 30.07 
psig, determine: 

(a) The refrigerating effect per pound. 

Ans. 48.18 Btu/lb 



88 PRINCIPLES OF REFRIGERATION 

(*) The weight of refrigerant circulated per (c) The volume of vapor to be displaced per 

minute per ton. Arts. 4.15 lb/min/ton minute per ton. Arts. 3.13 cuft/min/ton 

(c) The volume of vapor displaced per minute 4. K , in Problem 2, the liquid is subcooled to 

per ton. Arts. 3.77 cu ft/min/ton 70° F and the evaporating pressure is lowered to 

3. If the liquid approaching the refrigerant 16.35 psig, determine 

control in Problem 2 is subcooled to 70° F, (a) The refrigerating effect. Arts. 45.25 Btu/lb 

determine: (b) The weight of refrigerant circulated per 

(a) The refrigerating effect. Arts. 57.93 Btu/lb minute per ton. Arts. 4.42 lb/min/ton 

(b) The weight of refrigerant circulated per (c) The volume of vapor displaced per 
minute per ton. Arts. 3.45 cu ft/min/ton minute per ton. Arts. 6.44 cu ft/min/ton 



7 

Cycle Diagrams 
and the Simple 
Saturated Cycle 



7-1. Cycle Diagrams. A good knowledge of 
the vapor-compression cycle requires an inten- 
sive study not only of the individual processes 
that make up the cycle but also of the relation- 
ships that exist between the several processes and 
of the effects that changes in any one process in 
the cycle have on all the other processes in the 
cycle. This is greatly simplified by the use of 
charts and diagrams upon which the complete 
cycle may be shown graphically. Graphical 
representation of the refrigeration cycle permits 
the desired simultaneous consideration of all the 
various changes in the condition of the re- 
frigerant which occur during the cycle and the 
effect that these changes have on the cycle with- 
out the necessity of holding in mind all the 
different numerical values involved in cyclic 
problems. 

The diagrams frequently used in the analysis of 
the refrigeration cycle are the pressure-enthalpy 
(Ph) diagram and the temperature-entropy (Ts) 
diagram. Of the two, the pressure-enthalpy 
diagram seems to be the most useful and is the 
one which is emphasized in the following sec- 
tions. The temperature-entropy diagram has 
already been introduced (Section 4-19) and its 
application to the refrigeration cycle will be 
discussed to some extent in this chapter. 
7-2. The Pressure-Enthalpy Diagram. A 
pressure-enthalpy chart for R-12 is shown in 



Fig. 7-1.* The condition of the refrigerant in 
any thermodynamic state can be represented as 
a point on the Ph chart. The point on the Ph 
chart which represents the condition of the 
refrigerant in any one particular thermodynamic 
state may be located if any two properties of 
the refrigerant at that state are known. Once the 
state point has been located on the chart, all the 
other properties of the refrigerant for that state 
can be determined directly from the chart. 

As shown by the skeleton Ph chart in Fig. 
7-2, the chart is divided into three areas which 
are separated from each other by the saturated 
liquid and saturated vapor curves. The area 
on the chart to the left of the saturated liquid 
curve is called the subcooled region. At any 
point in the subcooled region the refrigerant is 
in the liquid state and its temperature is below 
the saturation temperature corresponding to 
its pressure. The area to the right of the satu- 
rated vapor curve is the superheated region and 
the refrigerant is in the form of a superheated 
vapor. The center section of the chart, between 
the saturated liquid and saturated vapor curves, 
represents the change in phase of the refrigerant 
between the liquid and vapor states. At any 
point between the two curves the refrigerant is 
in the form of a liquid-vapor mixture. The 
distance between the two curves along any 
constant pressure line, as read on the enthalpy 
scale at the bottom of the chart, is the latent heat 
of vaporization of the refrigerant at that 
pressure. The saturated liquid and saturated 
vapor curves are not exactly parallel to each 
other because the latent heat of vaporization 
of the refrigerant varies with the pressure at 
which the change in phase occurs. 

On the chart, the change in phase from the 
liquid to the vapor phase takes place progres- 
sively from left to right, whereas the change in 
phase from the vapor to the liquid phase occurs 
from right to left. Close to the saturated liquid 
curve the liquid-vapor mixture is nearly all 
liquid, whereas close to the saturated vapor 
curve the liquid-vapor mixture is almost all 
vapor. The lines of constant quality (Fig. 7-3), 
extending from top to bottom through the 
center section of the chart and approximately 
parallel to the saturated liquid and vapor 

* The pressure-enthalpy chart for each refrigerant 
is different, depending upon the properties of the 
particular refrigerant. 



89 



90 PRINCIPLES OF REFRIGERATION 




1 

2 

£ 



3 

o. 



(eisd) ajnsssjd 



CYCLE DIAGRAMS AND THE SIMPLE SATURATED CYCLE 91 



Subcooled region 
(Refrigerant is in 
the form of a 
subcooled liquid) 




Region of phase change 
(Refrigerant is a liquid- 
vapor mixture) 



>— Liquid to vapor > > 

Vapor to liquid — < «— 



Superheated region 
(Refrigerant is in 
the form of a 
superheated vapor) 



Saturated liquid curve 



Saturated vapor curve- 



Specific enthalpy (Btu per lb) 
Fig. 7-2. Skeleton Mi chart Illustrating the three regions of the chart and the direction of phase changing. 



curves, indicate the percentage of vapor in the 
mixture in increments of 10%. For example, at 
any point on the constant quality line closest to 
the saturated liquid curve the quality of the 
liquid-vapor mixture is 10%, which means that 
10% (by weight) of the mixture is vapor. 
Similarly, the indicated quality of the mixture 
at any point along the constant quality line 
closest to the saturated vapor curve is 90% and 
the amount of vapor in the liquid-vapor mixture 
is 90%. At any point on the saturated liquid 
curve the refrigerant is a saturated liquid and at 
any point along the saturated vapor curve the 
refrigerant is a saturated vapor. 

The pressure is plotted along the vertical 
axis, and the enthalpy is plotted along the 
horizontal axis. Hence, the horizontal lines 
extending across the chart are lines of constant 
pressure and the vertical lines are lines of 
constant enthalpy. 

The lines of constant temperature in the 
subcooled region are almost vertical on the 
chart and paralled to the lines of constant 
enthalpy. In the center section, since the 
refrigerant changes state at a constant tempera- 
ture and pressure, the lines of constant tempera- 
ture run horizontally across the chart and 
parallel to the lines of constant pressure. At the 
saturated vapor curve the lines of constant 
temperature change direction again and, in the 



superheated vapor region, fall off sharply 
toward the bottom of the chart. 

The straight lines which extend diagonally 
and almost vertically across the superheated 
vapor region are lines of constant entropy. The 
curved, nearly horizontal lines crossing the 
superheated vapor region are lines of constant 
volume. 

The values of any of the various properties 
of the refrigerant which are of importance in the 
refrigerating cycle may be read directly from 
the Ph chart at any point where the value of that 
particular property is significant to the process 
occurring at that point. To simplify the chart, 
the number of lines on the chart is kept to a 
minimum. For this reason, the value of those 
properties of the refrigerant which have no real 
significance at some points in the cycle are 
omitted from the chart at these points. For 
example, in the liquid region and in the region 
of phase change (center section) the values of 
entropy and volume are of no particular 
interest and are therefore omitted from the 
chart in these sections. 

Since the Ph chart is based on a 1 lb mass of 
the refrigerant, the volume given is the specific 
volume, the enthalpy is in Btu per pound, and 
the entropy is in Btu per pound per degree of 
absolute temperature. Enthalpy values are 
found on the horizontal scale at the bottom of 



92 PRINCIPLES OF REFRIGERATION 

the chart and the values of entropy and volume 
are given adjacent to the entropy and volume 
lines, respectively. The values of both enthalpy 
and entropy are based on the arbitrarily selected 
zero point of —40° F. 

The magnitude of the pressure in psia is 
read on the vertical scale at the left side of the 
chart. Temperature values in degrees Fahren- 
heit are found adjacent to the constant tempera- 
ture lines in the subcooled and superheated 
regions of the chart and on both the saturated 
liquid and saturated vapor curves. 
7-3. The Simple Saturated Refrigerating 
Cycle. A simple saturated refrigerating cycle is 
a theoretical cycle wherein it is assumed that the 
refrigerant vapor leaves the evaporator and 
enters the compressor as a saturated vapor (at 
the vaporizing temperature and pressure) and 
the liquid leaves the condenser and enters the 
refrigerant control as a saturated liquid (at 
the condensing temperature and pressure). 
Although the refrigerating cycle of an actual 
refrigerating machine will usually deviate 
somewhat from the simple saturated cycle, the 
analysis of a simple saturated cycle is nonethe- 
less worthwhile. In such a cycle, the funda- 
mental processes which are the basis of every 
actual vapor compression refrigerating cycle are 
easily identified and understood. Furthermore, 



by using the simple saturated cycle as a standard 
against which actual cycles may be compared, 
the relative efficiency of actual refrigerating 
cycles at various operating conditions can be 
readily determined. 

A simple saturated cycle for a R-12 system is 
plotted on a Ph chart in Fig. 7-4. The system is 
assumed -to be operating under such conditions 
that the vaporizing pressure in the evaporator is 
35.75 psia and the condensing pressure in the 
condenser is 131.6 psia. The points A, B, C, D, 
and E on the Ph diagram correspond to points 
in the refrigerating system as shown on the flow 
diagram in Fig. 7-5. 

At point A, the refrigerant is a saturated 
liquid in the condenser at the condensing 
pressure and temperature, and its properties, as 
given in Table 16-3, are: 

p = 131.6 psia t = 100° F 

h = 31.16 Btu/lb s = 0.06316 Btu/lb/° F 
v = 0.0127 cu ft/lb 

At point A, the values of p, t, and h may be 
read directly from the Ph chart. Since the 
refrigerant is always a saturated liquid at point 
A, point A will always fall somewhere along 
the saturated liquid curve and can be located on 
the Ph chart if either/?, t, or h is known. Usually 



25 




Specific enthalpy (Btu per lb) 



Fig. 7-3. Skeleton Ph chart showing paths of constant pressure, constant temperature, constant volume, 
constant quality, constant enthalpy, and constant entropy. (Refrigerant- 1 2.) 



CYCLE DIAGRAMS AND THE SIMPLE SATURATED CYCLE 93 




h x h a he h e hd 

Specific enthalpy (Btu per lb) 
Fig. 7-4. Pressure-enthalpy diagram of a simple saturated cycle operating at a vaporizing temperature of 20° F 
and a condensing temperature of 100° F. (Refrigerant- 1 2.) 



in actual practice, either p, t, or both will be 
measurable. 

7-4. The Expansion Process. In the simple 
saturated cycle there is assumed to be no change 
in the properties (condition) of the refrigerant 
liquid as it flows through the liquid line from 
the condenser to the refrigerant control and 
the condition of the liquid approaching the re- 
frigerant control is the same as its condition 
at point A. The process described by the initial 
and final state points A-B occurs in the refriger- 
ant control when the pressure of the liquid 
is reduced from the condensing pressure to the 
evaporating pressure as the liquidpassesthrough 
the control.* When the liquid is expanded into 
the evaporator through the orifice of the control, 
the temperature of the liquid is reduced from the 
condensing temperature to the evaporating tem- 
perature by the flashing into vapor of a small 
portion of the liquid. 
Process A-B is a throttling type of adiabatic 

* Process A-B is an irreversible adiabatic ex- 
pansion during which the refrigerant passes through 
a series of state points in such a way that there is no 
uniform distribution of any of the properties. 
Hence, no true path can be drawn for the process and 
line A-B merely represents a process which begins 
at state point A and terminates at state point B. 



expansion, frequently called "wire-drawing," 
in which the enthalpy of the working fluid does 
not change during the process. This type of 
expansion occurs whenever a fluid is expanded 
through an orifice from a high pressure to a 
lower pressure. It is assumed to take place 
without the gain or loss of heat through the 
piping or valves and without the performance of 

work.f 

Since the enthalpy of the refrigerant does not 
change during process A-B, point B is located 
on the Ph chart by following the line of constant 
enthalpy from point A to the point where the 
constant enthalpy line intersects the line of 
constant pressure corresponding to the evapor- 
ating pressure. To locate point B on the Ph 
chart, the evaporating pressure or temperature 
must be known. 

As a result of the partial vaporization of 
the liquid refrigerant during process A-B, the 

t Actually, a certain amount of work is done by 
the fluid in projecting itself through the orifice of 
the control. However, since the heat equivalent 
of the work done in overcoming the friction of the 
orifice merely heats the orifice and is subsequently 
reabsorbed by the fluid, the assumption that the 
enthalpy of the fluid does not change during the 
process is not in error. 



94 PRINCIPLES OF REFRIGERATION 



refrigerant at point B is a liquid-vapor mixture 
whose properties are : 
p = 35.75 psia 
t = 20° F 

h =31.16 Btu/lb (same as at point A) 
v = 0.1520 cu ft/lb 
s - 0.06316 Btu/lb/°F 

Note. The change in entropy during the 
process A-B results from a transfer of heat 
energy which takes place within the refrigerant 
itself because of internal friction. A transfer of 
energy which occurs entirely within the working 
fluid does not affect the enthalpy of the fluid, 
only the entropy changes. 

At point B, in addition to the values of p, t, 
and h, the approximate quality of the vapor can 
be determined from the Ph chart by interpolat- 
ing between the lines of constant quality. In 
this instance, the quality of the vapor as deter- 
mined from the Ph chart is approximately 
27%. 

Since the refrigerant at point B is a liquid- 
vapor mixture, only the values of p and t can be 
read directly from Table 16-3. However, 
because the enthalpy of the refrigerant at points 
A and B is the same, the enthalpy at point B 
may be read from Table 16-3 as the enthalpy 
at the conditions of point A. The quality of the 
vapor at point B can be determined as in Section 
6-22, using enthalpy values taken either from 
Table 16-3 or from the Ph chart directly. 



The values of s and v at point B are usually of 
no interest and are not given either on the Ph 
chart or in the vapor tables. If the values of s 
and v are desired, they must be calculated. 
7-5. The Vaporizing Process. The process 
B-C is the vaporization of the refrigerant in the 
evaporator. Since vaporization takes place at a 
constant temperature and pressure, B-C is both 
isothermal and isobaric. Therefore, point C 
is located on the Ph chart by following the lines 
of constant pressure and constant temperature 
from point B to the point where they intersect 
the saturated vapor curve. At point C the 
refrigerant is completely vaporized and is a 
saturated vapor at the vaporizing temperature 
and pressure. The properties of the refrigerant 
at point C, as given in Table 16-3 or as read 
from the Ph chart, are: 

p •* 35.75 psia (same as at point B) 

t = 20° F (same as at point B) 

h = 80.49 Btu/lb 

v =• 1.121 cu ft/lb 

s = 0.16949 Btu/lb/°F 
The enthalpy of the refrigerant increases 
during process B-C as the refrigerant flows 
through the evaporator and absorbs heat from 
the refrigerated space. The quantity of heat 
absorbed by the refrigerant in the evaporator 
(refrigerating effect) is the difference between 
the enthalpy of the refrigerant at points B and 
C. Thus, if h a , h b , h , h d , h e , and h x represent the 



c 

Point at which 
vaporization isv£~~ 
complete ^> 



Suction vapor flows 
from the evaporator 
to the compressor 
through the suction 
line without a 
change in condition 



Discharge vapor 
'from compressor 




Refrigerant after 
passing through 
refrigerant control 






Point at which 
condensation- 
begins 




In the simple saturated 
cycle, the refrigerant 
flows through the liquid 
line from the condenser 
to the refrigerant 
control without a 
change in condition 



/-S 



Z) 



£l 



Point at which r 
condensation is—' 
complete 



Fig. 7-5. Flow diagram of a 
simple saturated cycle. 



CYCLE DIAGRAMS AND THE SIMPLE SATURATED CYCLE 95 



enthalpies of the refrigerant at points A,B,C,D, 
E, and X, respectively, then 

qi=h c - K (7-1) 

where q x = the refrigerating effect in Btu/lb. 
But since h t is equal to h a , then 

^ = A. - K (7-2) 

When we substitute the appropriate values in 
Equation 7-2 for the example in question, 

ft = 80.49 - 31.16 

= 49.33 Btu/lb 

On the Ph diagram, the distance between 
point X and point C represents the total latent 
heat of vaporization of 1 lb of R-12 at the 
vaporizing pressure of 35.75 psia (h fg in Table 
16-3). Therefore, since the distance B-C is the 
useful refrigerating effect, the difference between 
X-C and B-C, which is the distance X-B, is 
the loss of refrigerating effect. 
7-6. The Compression Process. In the simple 
saturated cycle, the refrigerant undergoes no 
change in condition while flowing through the 
suction line from the evaporator to the com- 
pressor. Process C-D takes place in the 
compressor as the pressure of the vapor is 
increased by compression from the vaporizing 
pressure to the condensing pressure. For the 
simple saturated cycle, the compression process, 
C-D, is assumed to be isentropic* An isen- 
tropic compression is a special type of adiabatic 
process which takes place without friction, f 
It is sometimes described as a "frictionless- 
adiabatic" or "constant-entropy" compression. 

According to Equation 4-3, Section 4-19, the 
change in entropy (As) during any process is 
equal to the transferred heat (Ag) divided by the 
average absolute temperature ( C R). In any 
frictionless-adiabatic process, such as the com- 

* It will be shown later that compression of the 
vapor in an actual refrigerating compressor usually 
deviates somewhat from true isentropic compression. 
As a general rule, compression is polytropic. 

t The term, adiabatic, is used to describe any 
number of processes which take place without the 
transfer of energy as heat to or from the working 
substance during the process. Thus, an isentropic 
process is only one of a number of different processes 
which may be termed adiabatic. For example, 
compare process C-D with process A-B. Both are 
adiabatic, but C-D is frictionless, whereas A-B is 
a throttling type of process which involves friction. 



pression process C-D, wherein no heat, as such, 
is transferred either internally (within the vapor 
itself) or externally (to or from an external 
source) Ag will always be equal to zero. If Ag 
is equal to zero, then As must also be equal to 
zero. Hence, there is no change in the entropy 
of the vapor during a frictionless-adiabatic 
(isentropic) compression. 

Since there is no change in the entropy of the 
vapor during process C-D, the entropy of the 
refrigerant at point D is the same as at point & 
Therefore, point D can be located on the Ph 
chart by following the line of constant entropy 
from point C to the point where the constant 
entropy line intersects the hne of constant pres- 
sure corresponding to the condensing pressure. 

At point D, the refrigerant is a superheated 
vapor whose properties are : 

p = 131.6 psia 

t = 112° F (approximate) 

h = 90.6 Btu/lb (approximate) 

v = 0.330 cu ft/lb (approximate) 

s = 0.16949 Btu/lb/° F (same as at point C) 

All of the properties of the refrigerant at the 
condition of point D are taken from the Ph 
chart. Since the values of t, h, and v require 
interpolation, they are only approximations. 
The properties of the superheated refrigerant 
vapor cannot usually be read accurately from 
the vapor table unless the pressure of the vapor 
in question corresponds exactly to one of the 
pressures listed in the table, This is seldom the 
case, particularly at the higher pressures where 
the pressure listings in the table are in 10 lb 
increments. 

Work is done on the vapor during the com- 
pression process, C-D, and the enthalpy of the 
refrigerant is increased by an amount equal to 
the heat energy equivalent of the mechanical 
work done on the vapor. The heat energy 
equivalent of the work done during the com- 
pression is often referred to as the heat of 
compression and is equal to the difference in the 
enthalpy of the refrigerant at points D and C. 
Thus, where q t is the heat of compression per 
pound of refrigerant circulated, 

q t =h a - h e (7-3) 

For the example in question, 

q 2 = 90.60 - 80.49 
= 10.11 Btu/lb 



96 



PRINCIPLES OF REFRIGERATION 



The mechanical work done on the vapor 
by the piston during the compression may be 
calculated from the heat of compression. If 
h> is the work done in foot-pounds per pound of 
refrigerant circulated and / is the mechanical 
energy equivalent of heat, then 



w = ft x / 


(7-4) 


or w = J(h a - hj 


(7-5) 


when we substitute in Equation 7-4, 




w = 10.11 x 778 




- 7865.58 ft-lb 





As a result of absorbing the heat of com- 
pression, the hot vapor discharged from the 
compressor is in a superheated condition, that 
is, its temperature is greater than the saturation 
temperature corresponding to its pressure. 
In this instance, the vapor leaves the compressor 
at a temperature of 1 12° F, whereas the satura- 
tion temperature corresponding to its pressure 
of 131.6 psia is 100° F. Thus, before the vapor 
can be condensed, the superheat must be 
removed and the temperature of the vapor 
lowered from the discharge temperature to the 
saturation temperature corresponding to its 
pressure. * 

7-7. The Condensing Process. % Usually, both 
processes D-E and E-A take place in the 
condenser as the hot gas discharged from the 
compressor is cooled to the condensing tempera- 
ture and condensed. Process D-E occurs in the 
upper part of the condenser and to some extent 
in the hot gas line. It represents the cooling of 
the vapor from the discharge temperature to the 
condensing temperature as the vapor rejects 
heat to the condensing medium. During 
process D-E, the pressure of the vapor remains 
constant and point E is located on the Ph chart 
by following a line of constant pressure from 
point D to the point where the constant pressure 
line intersects die saturated vapor curve. 

At point E, the refrigerant is a saturated vapor 
at the condensing temperature and pressure. 
Its properties, as read from either the Ph chart 
or from Table 16-3, are: 

p = 131.6 psia (same as at point D) 
t = 100° F 
h = 88.62 Btu/lb 
s - 0.16584 Btu/lb/°F 
»» 0.319 cu ft/lb 



The quantity of sensible heat (superheat) 
removed from 1 lb of vapor in the condenser in 
cooling the vapor from the discharge tempera- 
ture to the condensing temperature is the differ- 
ence between the enthalpy of the refrigerant at 
point D and the enthalpy at point E (h a — h t ). 

Process E-A is the condensation of the vapor 
in the condenser. Since condensation takes 
place at a constant temperature and pressure, 
process E-A follows along lines of constant 
pressure and temperature from point E to 
point A. The heat rejected to the condensing 
medium during process E-A is the difference 
between the enthalpy of the refrigerant at 
points E and A (h e — h a ). 

On returning to point A, the refrigerant has 
completed one cycle and its properties are the 
same as those previously described for point A. 

Since both processes D-E and E-A occur in 
the condenser, the total amount of heat rejected 
by the refrigerant to the condensing medium in 
the condenser is the sum of the heat quantities 
rejected during processes D-E and D-A. The 
total heat given up by the refrigerant at the 
condenser is the difference between the enthalpy 
of the superheated vapor at point D and the 
saturated liquid at point A. Hence, 

?s = h ~ K (7-6) 

where ft = the heat rejected at the condenser 
per pound of refrigerant circulated. 
In this instance, 

ft = 90.60 - 31.16 
- 59.44 Btu/lb 

If the refrigerant is to reach point A at the 
end of the cycle in the same condition as it left 
point A at the beginning of the cycle, the total 
heat rejected by the refrigerant to the condens- 
ing medium in the condenser must be exactly 
equal to the heat absorbed by the refrigerant at 
all other points in the cycle. In a simple 
saturated cycle, the refrigerant is heated at only 
two points in the cycle: (1) in the evaporator 
by absorbing heat from the refrigerated space 
(ft) and in the compressor by the heat of com- 
pression (ft). 

Therefore, 

ft - ft + ft (7-7) 

In this instance, 

ft =49.33 +10.11 

= 59.44 Btu/lb 



CYCLE DIAGRAMS AND THE SIMPLE SATURATED CYCLE 97 



Where m is the weight of refrigerant to be 
circulated per minute per ton, 

200 Btu/min 



m 



1x 



(7-8) 



For the cycle in question, 
200 

= 4.05 lb/min/ton 

Then, where Q a is the total quantity of heat 
rejected at the condenser per minute per ton, 

Qa = »<fc) (7-9) 

or ft = mUhi ~ A„) (7-10) 

For the cycle in question, 

ft = 4.05 x 59.44 

= 240.93 Btu/min/ton 

Where Q % is the heat of compression per 
minute per ton of refrigerating capacity, 

e a =/w(? 8 ) (7-11) 

or ft = m(fi d - A,) (7-12) 

Substituting, 

ft =4.05 x 10.11 
= 40.95 Btu/min/ton 

Where W is the work of compression done on 
the vapor per minute per ton of refrigerating 
capacity, 

W = rriiw) (7-13) 

or, since w equals J(qJ or /(A„ - A„), 

W = JmUqd (7-14) 

or W = Jm(h d - h£ (7-15) 

or W = /(ft) (7-16) 

When we substitute in Equation 7-15, 

W = 778 x 4.05 x (90.60 - 80.49) 

= 31,856 ft-lb/min/ton 

7-8. Theoretical Horsepower. The theoreti- 
cal horsepower required to drive the compressor 
per ton of refrigerating capacity may be found 
by applying Equation 1-5 (Section 1-11): 

31,856 
P ~ 33,000 x 1 
= 0.965 hp/ton 



A more convenient method of determining 
the theoretical horsepower per ton is produced 
by combining Equations 1-5 and 7-15: 



hp = 



m(h d — Aj) 
42.42 



(7-17) 



The compressor horsepower as calculated 
above represents only the horsepower required 
to compress the vapor. That is, it is the 
theoretical power which would be required per 
ton of refrigerating capacity by a 100% efficient 
system. It does not take into account the power 
required to overcome friction in the compressor 
and other power losses. The actual (brake) 
horsepower required per ton of refrigeration 
will usually be from 30% to 50% more than 
the theoretical horsepower calculated, depend- 
ing upon the efficiency of the compressor. 
The factors governing compressor efficiency are 
discussed later. 

7-9. Coefficient of Performance. The coeffi- 
cient of performance of a refrigerating cycle is 
an expression of the cycle efficiency and is 
stated as the ratio of the heat absorbed in the 
refrigerated space to the heat energy equivalent 
of the energy supplied to the compressor, that is, 

Heat absorbed from 
Coefficient of _ the refrigerated space 
performance Heat energy equivalent 
of the energy supplied 
to the compressor 

For the theoretical simple saturated cycle, this 
may be written as 

Refrigerating effect 



c.o.p. = 

r Heat of compression 

(A, - kg) 

= (A d - A„) 

(ft) 
Hence, for the cycle in question, 
49.33 



(7-18) 



c.o.p. 



10.11 
4.88 



7-10. Effect of Suction Temperature on 
Cycle Efficiency. The efficiency of the vapor- 
compression refrigerating cycle varies consider- 
ably with both the vaporizing and condensing 



98 



PRINCIPLES OF REFRIGERATION 





h a h c h c ' h e hd hd 

Specific enthalpy (Btu/lb above - 40° F) 
Fig. 7-4. Comparison of two simple saturated cycles operating at different vaporizing temperatures (figure 
distorted). (Refrigerant- 12.) 



temperatures. Of the two, the vaporizing 
temperature has by far the greater effect. 

To illustrate the effect that varying the suction 
temperature has on cycle efficiency, cycle 
diagrams of two simple saturated cycles operat- 
ing at different suction temperatures are drawn 
on the Ph chart in Fig. 7-6. One cycle, identified 
by the points A, B, C, D, and E, is operating at a 
vaporizing temperature of 10° F and a condens- 
ing temperature of 100° F. A similar cycle 
having the same condensing temperature but 
operating at a vaporizing temperature of 40° F 
is set off by the points A, W, C", £>', and E. 

To facilitate a comparison of the two cycles, 
the following values have been determined from 
the PA chart: 

(a) For the 10° F cycle, 
qi=h c -h a = 79.36 - 31.16 = 48.20 Btu/lb 
q i =h d -h e = 90.90 - 79.36 = 11.54 Btu/lb 
q a "h d -h a = 90.90 - 31.16 = 59.74 Btu/lb 

(6) For the 40° F cycle, 
qi = h c . -h a = 82.71 - 31.16 = 51.55 Btu/lb 
q a = h* - A < = 90.20 - 82.71 = 7.49 Btu/lb 
fl 8 = h, -h a = 90.20 - 31.16 = 59.04 Btu/lb 

In comparing the two cycles, note that the 
refrigerating effect per pound of refrigerant 
circulated is greater for the cycle having the 
higher vaporizing temperature. The refrigerat- 



ing effect for the cycle having the 10° F vaporiz- 
ing temperature is 48.20 Btu/lb. When the 
vaporizing temperature of the cycle is raised to 
40° F, the refrigerating effect increases to 51.55 
Btu/lb. This represents an increase in the 
refrigerating effect per pound of 



(he - K) - {h c - h a ) 
h c — h a 

51.55 - 48.20 



x 100 



x 100 



48.20 
= 6.95% 

The greater refrigerating effect per pound of 
refrigerant circulated obtained at the higher 
vaporizing temperature is accounted for by the 
fact that there is a smaller temperature differen- 
tial between the vaporizing temperature and the 
temperature of the liquid approaching the 
refrigerant control. Hence, at the higher 
suction temperature, a smaller fraction of the 
refrigerant vaporizes in the control and a greater 
portion vaporizes in the evaporator and pro- 
duces useful cooling. 

Since the refrigerating effect per pound is 
greater, the weight of refrigerant which must be 
circulated per minute per ton of refrigerating 
capacity is less at the higher suction temperature 
than at the lower suction temperature. Whereas 



CYCLE DIAGRAMS AND THE SIMPLE SATURATED CYCLE 99 



the weight of refrigerant circulated per minute 
per ton for the 10° F cycle is 

200 



Ac — f>a 

200 
= 48.20 
= 4.151b/min 

The weight of refrigerant circulated per minute 
per ton for the 40° F cycle is only 

200 



200 
~ 51.55 
- 3.88 lb/min 

The decrease in the weight of refrigerant 
circulated at the higher suction temperature is 



4.15 - 3.88 



4.15 



x 100 



= 6.5% 

Since the difference between the vaporizing 
and condensing pressures is smaller for the 
cycle having the higher suction temperature, 
the work of compression per pound required to 
compress the vapor from the vaporizing 
pressure to the condensing pressure is less for 
the higher temperature cycle than for the lower 
temperature cycle. It follows then that the heat 
of compression per pound for the cycle having 
the higher vaporizing temperature is also less 
than that for the cycle having the lower vaporiz- 
ing temperature. The heat of compression per 
pound for the 10° F cycle is 11.54 Btu, whereas 
the heat of compression for the 40° F cycle is 
only 7.49 Btu. This represents a decrease in the 
heat of compression per pound of 

(A„ - h e ) - {h d . - h<) 



x 100 



Because both the work of compression per 
pound and the weight of refrigerant circulated 



h d — h e 


11.54 -7.49 


11.54 


-35.1% 



per minute per ton are less at the higher suction 
temperature, the work of compression per 
ton and therefore the theoretical horsepower 
required per ton will be smaller at the higher 
suction temperature. The theoretical horse- 
power required per ton of refrigerating capacity 
for the 10° F cycle is 



42.42 



4.15 x (90.90 - 79.36) 



42.42 



= 1.13 



For the 40° F cycle, the theoretical horsepower 
required per ton is 



"Kb* — hf) 
42.42 



3.88 x (90.20 - 82.71) 



42.42 



0.683 



In this instance, increasing the vaporizing 
temperature of the cycle from 10° F to 40° F 
reduces the theoretical horsepower required per 
ton by 

1.13 -0.683 

1.13 X 10 ° 

= 39.5% 

Later, when the efficiency of the compressor 
is taken into consideration, it will be shown that 
the difference in the actual horsepower required 
per ton at the various suction temperatures is 
even greater than that indicated by theoretical 
computations. 

Since the coefficient of performance is an 
index of the power required per unit of refriger- 
ating capacity and, as such, is an indication of 
cycle efficiency, the relative efficiency of the two 
cycles can be determined by comparing their 
coefficients of performance. The coefficient of 
performance for the 10° F cycle is 

h c — h a 



48.20 
= 11.54 

= 4.17 



100 PRINCIPLES OF REFRIGERATION 

and the coefficient of performance for the 40° F 
cycle is 

hf — h a 

h d > - h(f 
51.55 
~ 7.49 
= 6.88 

It is evident that the coefficient of perform- 
ance, and hence the efficiency of the cycle, 
improves considerably as the vaporizing tem- 
perature increases. In this instance, increasing 
the suction temperature from 10° F to 40° F 
increases the efficiency of the cycle by 



6.88 - 4.17 
4.17 



x 100 - 65% 



Although the difference in the weight of 
refrigerant which must be circulated per minute 
per ton of refrigerating capacity at the various 
suction temperatures is usually relatively small, 
the volume of vapor that the compressor must 
handle per minute per ton varies greatly with 
changes in the suction temperature. This is 
probably one of the most important factors 
influencing the capacity and efficiency of a 
vapor-compression refrigerating system and is 
the one which is the most likely to be overlooked 
by the student when studying cycle diagrams. 
The difference in the volume of vapor to be 
displaced per minute per ton at the various 
suction temperatures can be clearly illustrated 
by a comparison of the two cycles in question. 

For the 10° F cycle, the volume of vapor to be 
displaced per minute per ton is 

m(v) =4.15 x 1.351 = 5.6cuft 

whereas, at the 40° F suction temperature, the 
volume of vapor to be displaced per minute per 
ton is 

ntv) = 3.88 x 0.792 - 3.075 cu ft 

It is of interest to note that, whereas the 
decrease in the weight of refrigerant circulated 
per minute per ton at the higher suction tempera- 
ture is only 6.5%, the decrease in the volume 
of vapor handled by the compressor per minute 
per ton is 

5.6 - 3.075 

~^— x 100 =45% 



Obviously, then, the smaller weight of 
refrigerant circulated per minute per ton 
accounts for only a very small part of the 
reduction in the volume of vapor displaced per 
minute per ton at the higher suction tempera- 
ture. To a far greater extent, the decrease in 
the volume of vapor displaced per minute per 
ton is a result of the lower specific volume of the 
suction vapor which is coincident with the 
higher suction temperature (0.792 cu ft/lb at 
40° F as compared to 1.351 cu ft/lb at 10° F). 
This aspect of system capacity and efficiency in 
relation to suction temperature will be further 
investigated in conjunction with compressor 
performance in Chapter 12. 

The quantity of heat to be rejected at the 
condenser per minute per ton is much smaller 
for the cycle having the higher suction tempera- 
ture. This is true even though the quantity of 
heat rejected at the condenser per pound of 
refrigerant circulated is nearly the same for 
both cycles. For the 10° F cycle, the quantity 
of heat rejected at the condenser per minute per 
ton is 

Mh a -h tt ) = 4.15 x 59.74 = 247.92 

whereas for the 40° F cycle the heat rejected at 
the condenser per minute per ton is only 

"(Ad- - A„) = 3.88 x 59.04 = 229.08 Btu 
The quantity of heat rejected per minute per 
ton at the condenser is less for the higher 
suction temperature because of (1) the smaller 
weight of refrigerant circulated per minute per 
ton and (2) the smaller heat of compression per 
pound. 

It has been shown previously that the heat 
rejected at the condenser per pound of refriger- 
ant circulated is the sum of the heat absorbed 
in the evaporator per pound (refrigerating 
effect) and the heat of compression per pound. 
Since increasing the vaporizing temperature of 
the cycle brings about an increase in the 
refrigerating effect as well as a decrease in the 
heat of compression, the quantity of heat 
rejected at the condenser per pound remains 
very nearly the same for both cycles (59.74 at 
10° F as compared to 59.04 at 40° F). In 
general, this is true for all suction temperatures 
because any increase or decrease in the heat 
of compression is usually accompanied by an 
offsetting increase or decrease in the refrigerat- 
ing effect. 



CYCLE DIAGRAMS AND THE SIMPLE SATURATED CYCLE 



101 



171.8 
131.6 






1 



29.35 





x/** 


E'J D' 




a/\iw 


Eh Z)/^137.5'F 




















/X ! 

//l 1 ' 


/lO* 


B 


B 1 




// ' i ' 
fc l! ]! 






r i 
I 
I 


/ i 1 ! 1 

/i |S? I 
/ i i I i 




IS 
CO 


Si 


/tol cyl iir> ' |o 
/mi id. 1— | 'cm 

/ °> Sfi' Is 1 l< T > 

/ i II 1 1 



A a h a ' h c h e h e ' hi hi' 

Specific enthalpy (Btu/lb above - 40° F) 

Fig. 7-7. Comparison of two simple saturated cycles operating at different condensing temperatures (figure 
distorted). (Refrigerant- 1 2.) 



7-1 1. Effect of Condensing Temperature on 
Cycle Efficiency. Although the variations in 
cycle efficiency with changes in the condensing 
temperature are not as great as those brought 
about by changes in the vaporizing temperature, 
they are nonetheless important. In general, 
if the vaporizing temperature remains constant, 
the efficiency of the cycle decreases as the 
condensing temperature increases, and increases 
as the condensing temperature decreases. 

To illustrate the effect of condensing tempera- 
ture on cycle efficiency, cycle diagrams of two 
saturated cycles operating at different condens- 
ing temperatures are drawn on the Ph chart in 
Fig. 7-7. One cycle, A, B, C, D, and E, has a 
condensing temperature of 100° F, whereas the 
other cycle, A', B', C, D', and E', is operating at 
a condensing temperature of 120° F. The evapo- 
rating temperature of both cycles is 10° F. Values 
for cycle A-B-C-D-E have been determined in 
the previous section. Values for cycle A'-B'-C- 
D'-E' are as follows: 

From the Ph diagram, 

9! = h e - h a - = 79.36 - 36.16 = 43.20 Btu/lb 
qi = h d . - h e = 93.20 - 79.36 = 13.84 Btu/lb 
?s - h d . - h a . = 93.20 - 36.16 = 57.04 Btu/lb 

In a simple saturated cycle the liquid 
refrigerant reaches the refrigerant control at 



the condensing temperature. Therefore, as 
the condensing temperature is increased, the 
temperature of the liquid approaching the 
refrigerant control is increased and the refriger- 
ating effect per pound is reduced. In this 
instance, the refrigerating effect is reduced from 
48.20 Btu/lb to 43.20 Btu/lb when the condens- 
ing temperature is increased from 100° F to 
120° F. This is a reduction of 

48.20 - 43.20 

48.20 X 10 ° 

- 10.37% 

Because the refrigerating effect per pound is 
less for the cycle having the higher condensing 
temperature, the weight of refrigerant to be 
circulated per minute per ton must be greater. 
For the cycle having the 100° F condensing 
temperature the weight of refrigerant to be 
circulated per minute per ton is 4.15 lb. When 
the condensing temperature is increased to 
120° F, the weight of refrigerant which must be 
circulated per minute per ton increases to 

200 
4T20= 4 - 631b 



This is an increase of 
4.63 - 4.15 



4.15 



x 100 = 11.57% 



102 PRINCIPLES OF REFRIGERATION 



Since the weight of refrigerant which must be 
circulated per minute per ton is greater at the 
higher condensing temperature, it follows that 
the volume of vapor to be compressed per 
minute per ton must also be greater. In a simple 
saturated cycle the specific volume of the 
suction vapor varies only with the vaporizing 
temperature. As the vaporizing temperature 
is the same for both cycles, the specific volume 
of the vapor leaving the evaporator is also the 
same for both cycles and therefore the difference 
in the volume of vapor to be compressed per 
minute per ton is in direct proportion to the 
difference in the weight of refrigerant circulated 
per minute per ton. At the 100° F condensing 
temperature the volume of vapor to be com- 
pressed per minute per ton is 5.6 cu ft, whereas 
at the 120° F condensing temperature the 
volume of vapor compressed per minute per ton 
increases to 

4.63 x 1.351 =6.25cuft 

This represents an increase in the volume of 
vapor compressed per minute per ton of 

6.25 - 5.6 

— x 100 = 11.57% 

5.6 

Note that the percent increase in the volume 
of vapor handled by the compressor is exactly 
equal to the percent increase in the weight of 
refrigerant circulated. Contrast this with what 
occurs when the vaporizing temperature is 
varied. 

Since the difference between the vaporizing 
and condensing pressures is greater, the work 
of compression per pound of refrigerant 
circulated required to raise the pressure of the 
vapor from the vaporizing to the condensing 
pressure is also greater for the cycle having the 
higher condensing temperature. In this instance, 
the heat of compression increases from 11.54 
Btu/lb for the 100° F condensing temperature 
to 13.84 Btu/lb for the 120° F condensing 
temperature. This is an increase of 

13.84 - 11.54 

— iL54— Xl0 °= 20% 
As a result of the greater work of compression 
per pound and the greater weight of refrigerant 
circulated per minute per ton, the theoretical 
horsepower required per ton of refrigerating 
capacity increases as the condensing tempera- 
ture increases. Whereas the theoretical horse- 



power required per ton at the 100° F condensing 
temperature is only 1 . 1 3 hp when the condensing 
temperature is increased to 120° F, the theoreti- 
cal horsepower per ton increases to 

4.63 x 13.84 

42.42 =L52h P 
This is an increase in the power required per ton 

Note that the increase in the horsepower 
required per ton at the higher condensing 
temperature is greater than the increase in the 
work of compression per pound. This is 
accounted for by the fact that, in addition to the 
20% increase in the work of compression per 
pound, there is also a 6.5% increase in the 
weight of refrigerant circulated per minute per 
ton. 

The coefficient of performance of the cycle 
at the 100° F condensing temperature is 4.17. 
When the condensing temperature is raised to 
120° F, the coefficient of performance drops to 

43.20 

= 3.12 

13.84 

Since the coefficient of performance is an 
index of the refrigerating capacity per unit of 
power, the decrease in refrigerating capacity per 
unit of power in this instance is 

4.17 - 3.12 

— x 100 = 33.7% 

Obviously, the effect of raising the condensing 
temperature on cycle efficiency is the exact 
opposite of that of raising the evaporating 
temperature. Whereas raising the evaporating 
temperature increases the refrigerating effect per 
pound and reduces the work of compression so 
that the refrigerating capacity per unit of power 
increases, raising the condensing temperature 
reduces the refrigerating effect per pound and 
increases the work of compression so that the 
refrigerating capacity per unit of power de- 
creases. 

Although the quantity of heat rejected at the 
condenser per pound of refrigerant circulated 
varies only slightly with changes in the condens- 
ing temperature because any change in the heat 
of compression is accompanied by an offsetting 
change in the refrigerating effect per pound, 



CYCLE DIAGRAMS AND THE SIMPLE SATURATED CYCLE 



103 



Td 



Fig. 7-8. Temperature-entropy i 

diagram of simple saturated £ 

cycle on skeleton Ts chart M T * 

(figure distorted). (Refrigerant- J 

12.) I 




the total heat rejected at the condenser per 
minute per ton varies considerably with changes 
in the condensing temperature because of the 
difference in the weight of refrigerant circulated 
per minute per ton. It was shown in Section 7-7 
that the total heat rejected at the condenser per 
minute per ton (gj) is always the sum of the 
heat absorbed in the evaporator per minute per 
ton (Q x ) and the total heat of compression per 
ton (Q 2 ). Since Q x is always 200 Btu/min/ton, 
then Q 3 will vary only with Q 2 , the heat of 
compression per minute per ton. Furthermore, 
since Q 2 always increases as the condensing 
temperature increases, it follows then that Q 3 
also increases as .the condensing temperature 
increases. 

For the two cycles in question, the heat 
rejected at the condenser per minute per ton 
at the 100° F condensing temperature is 218.75 
Btu. For the 120° F condensing temperature, 
the quantity of heat rejected at the condenser per 
minute per ton increases to 

4.63 x (43.20 + 13.84) = 310.4 Btu 
The percent increase is 
310.40 - 218.75 
— lii™ >< 100 =41-8% 

It is interesting to note also that the amount 
of sensible heat rejected at the condenser in- 
creases considerably at the higher condensing 
temperature, whereas the amount of latent heat 
rejected diminishes slightly. This indicates that, 
at the higher condensing temperature, a 



Specific entropy (Btu/lb *F) 

greater portion of the condenser surface is being 
used merely to reduce the temperature of the 
discharge vapor to the condensing temperature. 
7-12. The Temperature-Entropy Diagram. 
Although the author is partial to the pressure- 
enthalpy diagram, there are many who prefer 
to use the temperature-entropy diagram to 
analyze the refrigeration cycle. To acquaint 
the student with the use of Ts diagrams in'cycle 
analysis, a diagram of the simple saturated cycle 
described in the foregoing sections is drawn 
on Ts coordinates in Fig. 7-8. The state points 
A, B, C, D, and E represent the points in the 
cycle as shown by the flow diagram in Fig. 7-5. 
The state point X represents saturated liquid 
at the vaporizing temperature. Process A-B is 
the irreversible adiabatic expansion through the 
refrigerant control.* Process B-C is the iso- 
baric and isothermal vaporization in the 
evaporator. Process C-D is the reversible 
adiabatic (isentropic) compression in the com- 
pressor and processes D-E and E-A are the 
desuperheating and condensing processes in the 
condenser. T t , T c , and T d are the absolute 
suction, absolute condensing, and absolute 
discharge temperatures, respectively, whereas 
s a , s b , s e , s d , s t , and s x are the specific entropies 
of the refrigerant at the various state points. 

The principal advantage of the Ts diagram is 
that the areas shown on the chart represent 
actual heat quantities. In Fig. 7-8, the area 

* The line A-B does not necessarily follow the 
actual path of process A-B. See Section 7-4. 



104 PRINCIPLES OF REFRIGERATION 

Condensing Temperature, 100° F 



Condensing Pressure, 136.16 Psia 



Suction temperature 


50° 


40° 


30° 


20° 


10° 


0° 


-10° 


-20° 


-30° 


-40° 


Absolute suction pressure 


61.39 


51.68 


43.16 


35.75 


29.35 


23.87 


19.20 


15.28 


12.02 


9.32 


Refrigerating effect per 
pound 


52.62 


51.55 


50.45 


49.33 


48.20 


47.05 


45.89 


44.71 


43.54 


42.34 


Weight of refrigerant 
circulated per minute per 
ton 


3.80 


3.88 


3.97 


4.05 


4.15 


4.25 


4.36 


4.48 


4.59 


4.73 


Specific volume of suction 
vapor 


0.673 


0.792 


0.939 


1.121 


1.351 


1.637 


2.00 


2.47 


3.09 


3.91 


Volume of vapor com- 
pressed per minute per ton 


2.56 


3.08 


3.77 


4.55 


5.60 


6.96 


8.72 


11.10 


14.20 


18.50 


Heat of compression per 
pound 


6.01 


7.49 


8.79 


10.11 


11.54 


13.29 


14.85 


16.73 


18.50 


20.40 


Theoretical horsepower per 
ton 


0.539 


0.683 


0.818 


0.965 


1.13 


1.35 


1.54 


1.78 


2.00 


2.26 


Coefficient of performance 


8.76 


6.88 


5.74 


4.88 


4.17 


3.54 


3.09 


2.67 


2.36 


2.07 



Fig. 7-9 




^40 -30 -20 -10 10 20 30 40 50 
Suction temperature 



4 
3.5 

3.0 a 

E 

2.5 f 
2-0 1 

Q. 

1.5 
1.0 
0.5 



20 
18 
16 

14 c 
12 - 
10 | 
8 * 
6 
4 
2 



Fig. 7-IOa. For Refrigerant- 1 2, the refrigerating effect per pound, the weight of refrigerant circulated per 
minute per ton, the specific volume of the suction vapor, and the volume of vapor compressed per minute 
per ton are each plotted against suction temperature. Condensing temperature is constant at 100° F. 



CYCLE DIAGRAMS AND THE SIMPLE SATURATED CYCLE 105 



Fig. 7-IOb. For Refrigerant- 1 2, 
the coefficient of performance 
and the horsepower per ton are 
plotted against suction tem- 
perature. Condensing tempera- 
ture is constant at 100° F. 



2.6 
2.2 



I 

o. 

Q) 

£ 
o 

I 



1.6 



1.2 



0.8 



0.4 



























\ 

V 






















\ 

V 






















^ 




















cop-P"* 










































^Horserjower 




















per ton 

1 











-40 - 

A-X-C-D-E-A represents the heat energy 
equivalent of the work of adiabatic compression 
and, since the distance between the base line and 
T„ foreshortened in the figure, represents the 
absolute vaporizing temperature of the liquid, 
area B-C-Sc-sg-B represents the refrigerating 
effect per pound. The sum of the areas A-X-C- 
D-E-A and B-C-Sg-sg-B, of course, represents 
the heat rejected at the condenser per pound. 
As in the case of the Ph diagram, it can be 
readily seen on the Ts chart that either lowering 
the vaporizing temperature or raising the con- 
densing temperature tends to increase the work 
of compression, reduce the refrigerating effect 
per pound, and lower the efficiency of the cycle. 
7-13. Summary. Regardless of the method 
used to analyze the cycle, it is evident that the 
capacity and efficiency of a refrigerating system 
improve as the vaporizing temperature increases 
and as the condensing temperature decreases. 
Obviously, then, a refrigerating system should 
always be designed to operate at the highest 
possible vaporizing temperature and the lowest 
possible condensing temperature commensurate 
with the requirements of the application. This 
will nearly always permit the most effective use 
of the smallest possible equipment and thereby 
effect a savings not only in the initial cost 
of the equipment but also in the operating 
expenses. 

In any event, the influence of the vaporizing 
and condensing temperatures on cycle efficiency 
is of sufficient importance to warrant a more 
intensive study. To aid the student in this, the 
relationship between the refrigerating effect per 



8 g 

a 

1 

■£ 
5 S 

i 

4$ 



30-20-10 10 20 30 40 50 
Suction temperature 

pound, the weight of refrigerant circulated per 
minute per ton, the specific volume of the suction 
vapor, the volume of vapor compressed per 
minute per ton, the horsepower required per 
ton, and the coefficient of performance of the 
cycle has been calculated for various suction 
temperatures. These values are given in tabular 
form in Fig. 7-9 and are illustrated graphically 
in Figs. 7-10a and 7-106. In addition, the effect 
of condensing temperature on the horsepower 
required per ton of refrigerating capacity is 
shown for several condensing temperatures in 
Fig. 7-11. 

Since the properties of the refrigerant at point 
D on the cycle diagram cannot ordinarily be 
obtained from the refrigerant tables and since 
these properties are difficult to read accurately 
from the Ph chart because of the size of the 
chart, the approximate isentropic discharge 



a 



1 



-40 -30 -20 -10 10 20 30 40 50 
Suction temperature 

Fig. 7-11. The effect of condensing temperature on 
the horsepower per ton. 





\ 






















v 


X 




















s 


\ 


v i 


*r1 


H„ 


















^ 


fcfc? 


*£ 

§ 


■>o 














sj 


fc^l 




-^_ 
















^"s; 


^jT 


-- 


— ~. 



























106 PRINCIPLES OF REFRIGERATION 



temperatures and the approximate enthalpy 
of the refrigerant vapor at point D have been 
compiled for a variety of vaporizing and con- 
densing temperatures and are given in Table 7-1 
to aid the student in arriving at more accurate 
solutions to the problems at the end of the 
chapter. 

PROBLEMS 

1. A Refrigerant-12 system operating on a 
simple saturated cycle has an evaporating tem- 
perature of 0° F and a condensing temperature 
of 110° F. Determine: 
A. (1) The refrigerating effect per pound of 
refrigerant circulated. 

Arts. 44.56 Btu/lb 
(2) The weight of refrigerant circulated 
per minute per ton. 

Arts. 4.49 lb/min/ton 



(3) The volume of vapor compressed per 
minute per ton. 

Ans. 7.35 cu ft/min/ton 

B. (1) The heat of compression per pound of 

refrigerant circulated. 

Ans. 14.39 Btu/lb 

(2) The heat of compression per minute 
per ton of refrigeration. 

Ans. 64.61 Btu/min/ton 

(3) The work of compression per minute 
per ton in foot-pounds. 

Ans. 50.267 lb/min/ton 

(4) The theoretical horsepower per ton. 

Ans. 1.52hp/ton 

(5) The coefficient of performance. 

Ans. 3.1 

C. (1) The heat rejected per minute per ton 

at the condenser. 

Ans. 264.61 Btu/min/ton 



8 

Actual 

Refrigerating 

Cycles 



8-1. Deviation from the Simple Saturated 
Cycle. Actual refrigerating cycles deviate some- 
what from the simple saturated cycle. The reason 
for this is that certain assumptions are made for 
the simple saturated cycle which do not hold 
true for actual cycles. For example, in the 
simple saturated cycle, the drop in pressure in 
the lines and across the evaporator, condenser, 
etc., resulting from the flow of the refrigerant 
through these parts is neglected. Furthermore, 
the effects of subcooling the liquid and of super- 
heating the suction vapor are not considered. 
Too, compression in the compressor is assumed 
to be true isentropic compression. In the 
following sections all these things are taken into 
account and their effect on the cycle is studied 
in detail.* 

8-2. The Effect of Superheating the Suction 
Vapor. In the simple saturated cycle, the suction 
vapor is assumed to reach the suction inlet of 
the compressor as a saturated vapor at the 
vaporizing temperature and pressure. In actual 
practice, this is rarely true. After the liquid 
refrigerant has completely vaporized in the 
evaporator, the cold, saturated vapor will 
usually continue to absorb heat and thereby 
become superheated before it reaches the 
compressor (Fig. 8-1). 

* The departure from true isentropic compression 
and the effect that it has on the cycle are discussed 
in Chapter 12. 



On the Ph diagram in Fig. 8-2, a simple 
saturated cycle is compared to one in which the 
suction vapor is superheated from 20° F to 
70° F. Points A , B, C, D, and E mark the satur- 
ated cycle, and points A, B, C", D', and E indi- 
cate the superheated cycle. 

If the slight pressure drop resulting from the 
flow of the vapor in the suction piping is 
neglected, it may be assumed that the pressure 
of the suction vapor remains constant during 
the superheating. That is, after the superheat- 
ing, the pressure of the vapor at the suction 
inlet of the compressor is still the same as 
the vaporizing pressure in the evaporator. 
With this assumption, point C'can be located on 
the Ph chart by following a line of constant 
pressure from point C to the point where the 
line of constant pressure intersects the 70° F 
constant temperature line. Point D' is found by 
following a line of constant entropy from point 
C to the line of constant pressure corresponding 
to the condensing pressure. 

In Fig. 8-2, the properties of the superheated 
vapor at points C" and D', as read from the Ph 
chart, are as follows: 

At point C", 

p = 35.75 psia / = 70° F 

v = 1 .260 cu ft/lb h = 88.6 Btu/lb 
s = 0.1840 Btu/lb/° R 



c 



, 20*F 



35.75 psia 



-*&■ 



c 



20* F 



35.75 psia 



D 



30"F C 20* F 

' 35.75 psia 4 \ 



50* F 



70*F 



J 



Saturated vapor 
■Superheated vapor 



8 



8 



^ XT 164* F_ 

35.75 r^i 131.6 psi'a 
psia 




100° E 



3 



100* F 



r 



c 



131.6 psia ") 



A 100* F 



/^\ 



131.6 psia 

Fig. 8-1. Flow diagram of superheated cycle. Liquid 
completely vaporized at point C — saturated vapor 
continues to absorb heat while flowing from C to 
C — vapor reaches compressor in superheated 
condition. Notice the high discharge temperature. 
(Refrigerant- 1 2.) 



107 



108 PRINCIPLES OF REFRIGERATION 



131.16 



a 



£ 35.75 




Superheat 




Enthalpy (Btu/lb) 
Fig. 8-2. Ph diagram comparing simple saturated cycle to the superheated cycle. (Refrigerant- 1 2). 



At point/)', 
/> = 131.6 psia f = 164°F 

v - 0.380 cu ft/lb h = 99.2 Btu/lb 
j = 0.1840 Btu/lb/° R 

On the Ph chart, process C-C represents the 
superheating of the suction vapor from 20° F 
to 70° F at the vaporizing pressure, and the 
difference between the enthalpy of the vapor 
at these points is the amount of heat required to 
superheat each pound of refrigerant. In 
comparing the two cycles, the following 
observations are of interest: 

1 . The heat of compression per pound for the 
superheated cycle is slightly greater than that 
for the saturated cycle. For the superheated 
cycle, the heat of compression is 

h x - h & = 99.2 - 88.6 = 10.6 Btu/lb 

whereas for the saturated cycle the heat of 
compression is 

h d -h e = 90.6 - 80.49 = 10.11 Btu/lb 

In this instance, the heat of compression per 
pound is greater for the superheated cycle by 

10.6 - 10.11 



10.11 



x 100 =» 4.84% 



2. For the same condensing temperature and 
pressure, the temperature of the discharge vapor 
leaving the head of the compressor is consider- 



ably higher for the superheated cycle than for 
the saturated cycle — in this case, 164° F for the 
superheated cycle as compared to 1 12° F for the 
saturated cycle. 

3. For the superheated cycle, a greater 
quantity of heat must be dissipated at the 
condenser per pound than for the saturated 
cycle. This is because of the additional heat 
absorbed by the vapor in becoming superheated 
and because of the small increase in the heat of 
compression per pound. For the superheated 
cycle, the heat dissipated at the condenser per 
pound is 

h- ~ K - 99.2 - 31.16 = 68.04 Btu/lb 

and for the saturated cycle the heat dissipated 
at the condenser per pound is 

h d -h a = 90.6 - 31.16 = 59.44 Btu/lb 

The percent increase in the heat dissipated at 

the condenser per pound for the superheated 

cycle is 

68.04-59.44 tnn ijjd , 
— x 100 = 14.4% 

Note that the additional heat which must be 
dissipated per pound at the condenser in the 
superheated cycle is all sensible heat. The 
amount of latent heat dissipated per pound is 
the same for both cycles. This means that in the 
superheated cycle a greater amount of sensible 
heat must be given up to the condensing medium 



ACTUAL REFRIGERATING CYCLES 



109 



before condensation begins and that a greater 
portion of the condenser will be used in cooling 
the discharge vapor to its saturation temperature. 
Notice also that, since the pressure of the 
suction vapor remains constant during the 
superheating, the volume of the vapor increases 
with the temperature approximately in accord- 
ance with Charles' law.* Therefore, a pound 
of superheated vapor will always occupy a 
greater volume than a pound of saturated vapor 
at the same pressure. For example, in Fig. 8-2, 
the specific volume of the suction vapor increases 
from 1.121 cu ft per pound at saturation to 
1 .260 cu ft per pound when superheated to 70° F. 
This means that for each pound of refrigerant 
circulated, the compressor must compress a 
greater volume of vapor if the vapor is super- 
heated than if the vapor is saturated. For this 
reason, in every instance where the vapor is 
allowed to become superheated before it reaches 
the compressor, the weight of refrigerant cir- 
culated by a compressor of any given displace- 
ment will always be less than when the suction 
vapor reaches the compressor in a saturated 
condition, provided the pressure is the same. 

The effect that superheating of the suction 
vapor has on the capacity of the system and on 
the coefficient of performance depends entirely 
upon where and how the superheating of the 
vapor occurs and upon whether or not the 
heat absorbed by the vapor in becoming super- 
heated produces useful cooling.* 
8-3. Superheating without Useful Cooling. 
Assume first that the superheating of the 
suction vapor occurs in such a way that no use- 
ful cooling results. When this is true, the refri- 
gerating effect per pound of refrigerant circulated 
is the same for the superheated cycle as for a 
saturated cycle operating at the same vaporizing 
and condensing temperatures, and therefore the 
weight of refrigerant circulated per minute per 
ton will also be the same for both the superheated 
and saturated cycles. Then, for both cycles 
illustrated in Fig. 8-2, 

* The temperature and volume of the vapor do 
not vary exactly in accordance with Charles' law 
because the refrigerant vapor is not a perfect gas. 

* The effects of superheating depend also upon 
the refrigerant used. The discussion in this chapter 
is limited to systems using R-12. The effects of 
superheating on systems using other refrigerants are 
discussed later. 



The weight of refrig- 


200 


erant circulated per 


fie ~ K 


minute per ton m 


200 




49.33 




= 4.05 lb/min/ton 



Since the weight of refrigerant circulated is the 
same for both the superheated and saturated 
cycles and since the specific volume of the vapor 
at the compressor inlet is greater for the super- 
heated cycle than for the saturated cycle, it 
follows that the volume of vapor that the com- 
pressor must handle per minute per ton of 
refrigerating capacity is greater for the super- 
heated cycle than for the saturated cycle. 



For the saturated 
cycle, the specific 
volume of the suction 
vapor v e 

The volume of 
vapor compressed per 
minute per ton V 

For the superheated 
cycle, the specific 
volume of the suction 
vapor «v 

The volume of vapor 
compressed per minute 
per ton V 



= 1.121 cu ft/lb 

= m x v 

= 4.05 x 1.121 

= 4.55 cu ft/min/ton 



= 1.260 cu ft/lb 

= m x v 

= 4.05 x 1.260 

= 5.02 cu ft/min/ton 



In regard to percentage, the increase in the 
volume of vapor which must be handled by a 
compressor operating on the superheated cycle 

This means, of course, that a compressor 
operating on the superheated cycle must be 
10.3% larger than the one required for the 
saturated cycle. 

Again, since the weight of refrigerant circu- 
lated per minute per ton is the same for both 
cycles and since the heat of compression per 
pound is greater for the superheated cycle than 
for the saturated cycle, it is evident that the 
horsepower per ton is greater for the super- 
heated cycle and the coefficient of performance 
is less. 



110 PRINCIPLES OF REFRIGERATION 



For the saturated cycle, 


m(h d - h e ) 


the horsepower per ton 


42.42 




4.05 x 10.11 




42.42 




= 0.965 hp/ton 


The coefficient of per- 


h e — h a 


formance 


h d -K 




49.33 




10.11 




= 4.88 


For the superheated 


m(h d - - h C ') 


cycle, the horsepower per 


42.42 


ton 


4.05 x 10.6 




42.42 




= 1.01 hp/ton 


The coefficient of per- 


h c ~h a 


formance 


h d ' - K- 




49.33 




10.60 




= 4.65 



In summary, when superheating of the vapor 
occurs without producing useful cooling, the 
volume of vapor compressed per minute per ton, 
the horsepower per ton, and the quantity of heat 
given up in the condenser per minute per ton are 
all greater for the superheated cycle than for the 
saturated cycle. This means that the com- 
pressor, the compressor driver, and the conden- 
ser must all be larger for the superheated cycle 
than for the saturated cycle. 
8-4. Superheating That Produces Useful 
Cooling. Assume, now, that all of the heat 
taken in by the suction vapor produces useful 
cooling. When this is true, the refrigerating effect 
per pound is greater by an amount equal to the 
amount of superheat. In Fig. 8-2, assuming that 
the superheating produces useful cooling, the 
refrigerating effect per pound for the superheated 
cycle is equal to 

h^ -h a = 88.60 - 31.16 = 57.44 Btu/lb 
Since the refrigerating effect per pound is 
greater for the superheated cycle than for the 
saturated cycle, the weight of refrigerant circu- 
lated per minute per ton is less for the super- 
heated cycle than for the saturated cycle. 
Whereas the weight of refrigerant circulated per 
minute per ton for the saturated cycle is 4.05, the 



weight of refrigerant circulated per minute per 
ton for the superheated cycle is 

200 200 

= 3.48 lb/min/ton 



h- - h n 



57.44 



Notice that, even though the specific volume 
of the suction vapor and the heat of compression 
per pound are both greater for the superheated 
than for the saturated cycle, the volume of vapor 
compressed per minute per ton and the horse- 
power per ton are less for the superheated cycle 
than for the saturated cycle. This is because of 
the reduction in the weight of refrigerant circu- 
lated. The volume of vapor compressed per 
minute per ton and the horsepower per ton for 
the saturated cycle are 4.55 cu ft and 0.965 hp, 
respectively, whereas for the superheated cycle 

The volume of = m x v c - 

vapor compressed per =3.48 x 1.260 
minute per ton V = 4.38 cu ft/min/ton 

The horsepower per m{h a > — A c .) 

ton 42.42 

3.48 x 10.60 



42.42 
= 0.870 hp/ton 

For the superheated cycle, both the refrig- 
erating effect per pound and the heat of com- 
pression per pound are greater than for the 
saturated cycle. However, since the increase in 
the refrigerating effect is greater proportionally 
than the increase in the heat of compression, the 
coefficient of performance for the superheated 
cycle is higher than that of the saturated cycle. 
For the saturated cycle, the coefficient of per- 
formance is 4.69, whereas for the superheated 
cycle 



The coefficient 
of performance 



(h c . - h a ) 57.44 



= 5.42 



(4r - h c ) 10.60 
It will be shown in the following sections that 
the superheating of the suction vapor in an 
actual cycle usually occurs in such a way that a 
part of the heat taken in by the vapor in becom- 
ing superheated is absorbed from the refrig- 
erated space and produces useful cooling, 
whereas another part is absorbed by the vapor 
after the vapor leaves the refrigerated space and 
therefore produces no useful cooling. The por- 
tion of the superheat which produces useful 
cooling will depend upon the individual applica- 
tion, and the effect of the superheating on the 



ACTUAL REFRIGERATING CYCLES 



cycle will vary approximately in proportion to 
the useful cooling accomplished. 

Regardless of the effect on capacity, except 
in some few special cases, a certain amount of 
superheating is nearly always necessary and, in 
most cases, desirable. When the suction vapor 
is drawn directly from the evaporator into the 
suction inlet of the compressor without at least 
a small amount of superheating, there is a good 
possibility that small particles of unvaporized 
liquid will be entrained in the vapor. Such a 
vapor is called a "wet" vapor. It will be shown 
later that "wet" suction vapor drawn into the 
cylinder of the compressor adversely affects the 
capacity of the compressor. Furthermore, since 
refrigeration compressors are designed as vapor 
pumps, if any appreciable amount of un- 
vaporized liquid is allowed to enter the com- 
pressor from the suction line, serious mechanical 
damage to the compressor may result. Since 
superheating the suction vapor eliminates the 
possibility of "wet" suction vapor reaching the 
compressor inlet, a certain amount of super- 
heating is usually desirable. Again, the extent 
to which the suction vapor should be allowed to 
become superheated in any particular instance 
depends upon where and how the superheating 
occurs and upon the refrigerant used. 

Superheating of the suction vapor may take 
place in any one or in any combination of the 
following places: 

1. In the end of the evaporator 

2. In the suction piping installed inside the 
refrigerated space (usually referred to as a "drier 
loop") 

3. In the suction piping located outside of the 
refrigerated space 

4. In a liquid-suction heat exchanger. 

8-5. Superheating in Suction Piping outside 
the Refrigerated Space. When the cool re- 
frigerant vapor from the evaporator is allowed 
to become superheated while flowing through 
suction piping located outside of the refrigerated 
space, the heat taken in by the vapor is absorbed 
from the surrounding air and no useful cooling 
results. It has already been demonstrated that 
superheating of the suction vapor which pro- 
duces no useful cooling adversely affects the 
efficiency of the cycle. Obviously, then, super- 
heating of the vapor in the suction line outside 



of the refrigerated space should be eliminated 
whenever practical. 

Superheating of the suction vapor in the 
suction line can be prevented for the most part 
by insulating the suction line. Whether or not 
the loss of cycle efficiency in any particular 
application is sufficient to warrant the additional 
expense of insulating the suction line depends 
primarily on the size of the system and on the 
operating suction temperature. 

When the suction temperature is relatively 
high (35° F or 40° F), the amount of superheating 
will usually be small and the effect on the 
efficiency of the cycle will be negligible. The 
reverse is true, however, when the suction tem- 
perature is low. The amount of superheating is 
apt to be quite large. 

Too, at low suction temperatures, when the 
efficiency of the cycle is already very low, each 
degree of superheat will cause a greater reduc- 
tion in cycle efficiency percentagewise than when 
the suction temperature is high. It becomes 
immediately apparent that any appreciable 
amount of superheating in the suction line of 
systems operating at low suction temperatures 
will seriously reduce the efficiency of the cycle 
and that, under these conditions, insulating of 
the suction line is not only desirable but 
absolutely necessary if the efficiency of the cycle 
is to be maintained at a reasonable level. 

Aside from any considerations of capacity, 
even at the higher suction temperatures, insulat- 
ing of the suction line is often required to prevent 
frosting or sweating of the suction line. In 
flowing through the suction piping, the cold 
suction vapor will usually lower the temperature 
of the piping below the dew point temperature 
of the surrounding air so that moisture will 
condense out of the air onto the surface of the 
piping, causing the suction piping to either frost 
or sweat, depending upon whether or not the 
temperature of the piping is below the freezing 
temperature of water. In any event, frosting or 
sweating of the suction piping is usually un- 
desirable and should be eliminated by insulating 
the piping. 

8-6. Superheating the Vapor inside the Re- 
frigerated Space. Superheating of the suction 
vapor inside the refrigerated space can take 
place either in the end of the evaporator or in 
suction piping located inside the refrigerated 
space, or both. 



112 PRINCIPLES OF REFRIGERATION 




Fig. 8-3. Flow diagram showing drier loop for super- 
heating suction vapor inside refrigerated space. 



To assure the proper operation of the refri- 
gerant control and to prevent liquid refrigerant 
from overflowing the evaporator and being 
carried back to the compressor, when certain 
types of refrigerant controls are used, it is 
necessary to adjust the control so that the liquid 
is completely evaporated before it reaches the 
end of the evaporator. In such cases, the cold 
vapor will continue to absorb heat and become 
superheated as it flows through the latter portion 
of the evaporator. Since the heat to superheat 
the vapor is drawn from the refrigerated space, 
useful cooling results and the refrigerating effect 
of each pound of refrigerant is increased by an 
amount equal to the amount of heat absorbed in 
the superheating. 

It has been shown that when the superheating 
of the suction vapor produces useful cooling the 
efficiency of the cycle is improved somewhat.* 
However, in spite of the increase in cycle effi- 
ciency, it must be emphasized that superheating 
the suction vapor in the evaporator is not 
economical and should always be limited to only 
that amount which is necessary to the proper 
operation of the refrigerant control. Since the 
transfer of heat through the walls of the evapora- 
tor per degree of temperature difference is not as 

* Although this is true for systems using R-12 as 
a refrigerant, it will be shown later that this is not 
true for all refrigerants. 



great to a vapor as to a liquid, the capacity of the 
evaporator is always reduced in any portion of 
the evaporator where only vapor exists. There- 
fore, excessive superheating of the suction vapor 
in the evaporator will reduce the capacity of the 
evaporator unnecessarily and will require either 
that the evaporator be operated at a lower 
vaporizing temperature or that a larger evapora- 
tor be used in order to provide the desired 
evaporator capacity. Neither of these is desir- 
able nor practical. Since the space available for 
evaporator installation is often limited and since 
evaporator surface is expensive, the use of a 
larger evaporator is not practical. Because of 
the effect on cycle efficiency, the undesirability of 
lowering the vaporizing temperature is obvious. 
Often, a certain amount of suction piping, 
usually called a drier loop, is installed inside the 
refrigerated space for the express purpose of 
superheating the suction vapor (Fig. 8-3). Use 
of a drier loop permits more complete flooding 
of the evaporator with liquid refrigerant without 
the danger of the liquid overflowing into the 
suction line and being drawn into the compres- 
sor. This not only provides a means of super- 
heating the suction vapor inside the refrigerated 
space so that the efficiency of the cycle is 
increased without the sacrifice of expensive eva- 
porator surface, but it actually makes possible 
more effective use of the existing evaporator 
surface. Also, in some instances, particularly 
where the suction temperature is high and the 
relative humidity of the outside air is reasonably 
low, superheating of the suction vapor inside the 
refrigerated space will raise the temperature of 
the suction piping and prevent the formation of 
moisture, thereby eliminating the need for suction 
line insulation. It should be noted, however, 
that the extent to which the suction vapor can be 
superheated inside the refrigerated space is 
limited by the space temperature. Ordinarily, if 
sufficient piping is used, the suction vapor can 
be heated to within 4° F to 5° F of the space 
temperature. Thus, for a 40° F space tempera- 
ture, the suction vapor may be superheated to 
approximately 35° F. 

8-7. The Effects of Subcooling the Liquid. 
On the PA diagram in Fig. 8-4, a simple saturated 
cycle is compared to one in which the liquid is 
subcooled from 100° F to 80° F before it reaches 
the refrigerant control. Points A, B, C, D, and E 
designate the simple saturated cycle, whereas 



points A', ff, C, D, and E designate the sub- 
cooled cycle. 

It has been shown (Section 6-28) that when 
the liquid is subcooled before it reaches the 
refrigerant control the refrigerating effect per 
pound is increased. In Fig. 8-4, the increase in 
the refrigerating effect per pound resulting from 
the subcooling is the difference between h h and 
h v , and is exactly equal to the difference between 
h a and h a ; which represents the heat removed 
from the liquid per pound during the subcooling. 



For the saturated cycle, 
the refrigerating effect per 
pound, q x 

For the subcooled cycle, 
the refrigerating effect per 
pound, q t 



— K — K 

= 80.49 -31.16 
= 49.33 Btu/lb 
= h e — h tt > 
= 80.49 - 26.28 
= 54.21 Btu/lb 



200 
~ 49.33 
= 4.05 lb 

200 



Because of the greater refrigerating effect per 
pound, the weight of refrigerant circulated per 
minute per ton is less for the subcooled cycle 
than for the saturated cycle. 

For the saturated cycle, the 
weight of refrigerant circulated 
per minute per ton m 

For the subcooled cycle, the 
weight of refrigerant circulated — 54.21 
per minute per ton m = 3.69 lb 

Notice that the condition of the refrigerant 
vapor entering the suction inlet of the compressor 
is the same for both cycles. For this reason, the 
specific volume of the vapor entering the com- 
pressor will be the same for both the saturated 
and subcooled cycles and, since the weight of 
refrigerant circulated per minute per ton is less 



ACTUAL REFRIGERATING CYCLES 113 

for the subcooled cycle than for the saturated 
cycle, it follows that the volume of vapor which 
the compressor must handle per minute per ton 
will also be less for the subcooled cycle than for 
the saturated cycle. 



For the saturated cycle, 
the specific volume of the 
suction vapor v e 

The volume of vapor 
compressed per minute per 
ton V 

For the subcooled cycle, 
the specific volume of the 
suction vapor v e 

The volume of vapor 
compressed per minute per 
ton V 



1.121 cu ft/lb 
m x v e 
4.05 x 1.121 
■■ 4.55 cu ft/min 



■ 1.121 cu ft/lb 

m x v e 

3.69 x 1.121 
: 4.15 cu ft/min 



Because the volume of vapor compressed per 
minute per ton is less for the subcooled cycle, 
the compressor displacement required for the 
subcooled cycle is less than that required for the 
saturated cycle. 

Notice also that the heat of compression per 
pound and therefore the work of compression 
per pound is the same for both the saturated and 
subcooled cycles. This means that the refri- 
gerating effect per pound resulting from the 
subcooling is accomplished without increasing 
the energy input to the compressor. Any change 
in the refrigerating cycle which increases the 
quantity of heat absorbed in the refrigerated 
space without causing an increase in the energy 
input to the compressor will increase the c.o.p. 
of the cycle and reduce the horsepower required 
per ton. 



Fig. 8-4. Ph diagrams compar- 
ing the subcooled cycle to 
the simple saturated cycle. 
(Refrigerant- 1 2.) 



131.16 



S 



: 35.75 





Enthalpy (Btu/lb) 



114 PRINCIPLES OF REFRIGERATION 



Liquid-vapor mixture 
35.75 psia-20°F 



m< 



K 



Superheated vapor 
131.16 paa-112'F 



Saturated vapor 
3575psia-20*F 




Saturated vapor 
131.16 psia-lOOV 



Subcooled liquid 
131.16 psia-SCF 



Liquid subcooled 20* 
by giving off heat to 
surrounding air while 
passing through 
liquid line, receiver, 
etc. , 



Fig. 8-5. Flow diagram illus- 
trating subcooling of the liquid 
in the liquid line. (Refrigerant- 
12.) 



) 



C 



K 



For the saturated cycle, 
the coefficient of perfor- 
mance 



The horsepower per ton 



For the subcooled cycle, 
the coefficient of perfor- 
mance 



h d - K 
80.49 - 


- 26.28 


90.60 - 
54.21 
10.51 
= 5.16 
m{h d - 


• 80.49 
h e ) 


42.42 
3.69 x 10.51 



Saturated liquid 
131.16 psia-100 o F 



K - K 

hd — h c 

80.49 - 31.16 
~ 90.60 - 80.49 
= 4.88 
_ m(h d — h c ) 

42A2 
= 4.05 x 10.51 
= 0.965 hp/ton 

h c . — h„' 



The horsepower per ton 



42.42 
= 0.914 hp/ton 

In this instance, the c.o.p. of the subcooled cycle 
is greater than that of the saturated cycle by 

5.H5 -4.88 

~^M~ X 10 ° = 5 - 7% 
Subcooling of the liquid refrigerant can and 
does occur in several places and in several ways. 
Very often the liquid refrigerant becomes sub- 
cooled while stored in the liquid receiver tank 
or while passing through the liquid line by giving 
off heat to the surrounding air (Fig. 8-5). In 
some cases where water is used as the condensing 
medium, a special liquid .subcooler is used to 
subcool the liquid (Fig. 8-6). The gain in system 
capacity and efficiency resulting from the liquid 



subcooling is very often more than sufficient to 
offset the additional cost of the subcooler, 
particularly for low temperature applications. 
The liquid subcooler may be piped either in 
series or in parallel with the condenser. When 
the subcooler is piped in series with the con- 
denser, the cooling water passes through the 
subcooler first and then through the condenser, 
thereby bringing the coldest water into contact 
with the liquid being subcooled (Fig. 8-7). 
There is some doubt about the value of a sub- 
cooler piped in series with the condenser. Since 
the cooling water is warmed by the heat absorbed 
in the subcooler, it reaches the condenser at a 
higher temperature and the condensing tempera- 
ture of the cycle is increased . Hence the increase 
in system efficiency resulting from the subcooling 
is offset to some extent by the rise in the con- 
densing temperature. 

When the subcooler is piped in parallel with 
the condenser (Fig. 8-6), the temperature of the 
water reaching the condenser is not affected by 
the subcooler. However, for either series or 
parallel piping, the size of the condenser water 
pump must be increased when a subcooler is 
added. If this is not done, the quantity of water 
circulated through the condenser will be 
diminished by the addition of the subcooler and 
the condensing temperature of the cycle will be 
increased, thus nullifying any benefit accruing 
from the subcooling. 

Notice that in each case discussed so far, 
the heat given up by the liquid in becoming 
subcooled is given up to some medium external 
to the system. 

8-8. Liquid-Suction Heat Exchangers. An- 
other method of subcooling the liquid is to bring 
about an exchange of heat between the liquid 



and the cold suction vapor going back to the 
compressor. In a liquid-suction heat exchanger, 
the cold suction vapor is piped through the 
heat exchanger in counterflow to the warm 
liquid refrigerant flowing through the liquid 
line to the refrigerant control (Fig. 8-8). In 
flowing through the heat exchanger the cold 
suction vapor absorbs heat from the warm 
liquid so that the liquid is subcooled as the 
vapor is superheated, and, since the heat 
absorbed by the vapor in becoming superheated 
is drawn from the liquid, the heat of the liquid 
is diminished by an amount equal to the amount 
of heat taken in by the vapor. In each of the 
methods of subcooling discussed thus far, the 
heat given up by the liquid in becoming sub- 
cooled is given up to some medium external to 
the system and the heat then leaves the system. 
When a liquid-suction heat exchanger is used, 
the heat given up by the liquid in becoming 
subcooled is absorbed by the suction vapor and 
remains in the system. 

On the Ph diagram in Fig. 8-9, a simple 
saturated cycle is compared to one in which a 
liquid-suction heat exchanger is employed. 
Points A, B, C, Z),and £ identify the saturated 
cycle and points A', B', C", D', E identify the 
in which the heat exchanger is used. In the cycle 
latter cycle, it is assumed that the suction vapor 
is superheated from 20° F to 60° F in the heat 
exchanger. 

The heat absorbed per pound of vapor in the 
heat exchanger is 

h c . -h c = 86.20 - 80.49 = 5.71 Btu/lb 







D 5 

7 

S3 



Water from subcooler 



Water-cooled 

condenser 

(10O*F condensing) 



"^ 



t/80- 



100* 



75' » 



i tower 



or city main 



Liquid to 
'subcooler 



90* water to cooling tower or sewer 

Fig. 8-4. Flow diagram illustrating subcooler piped 
in series with condenser. 



ACTUAL REFRIGERATING CYCLES 



115 




Water from^/ 
condenser 



Saturated liquid 
to subcooler 



Fig. 8-7. Flow diagram showing parallel piping for 
condenser and subcooler. 



Since the heat given up by the liquid in 
the heat exchanger in becoming subcooled is 
exactly equal to the heat absorbed by the vapor 
in becoming superheated, h a — h a - is equal to 
h C ' — h c and therefore is also equal to 5.71 
Btu/lb. Since h a — h a ' represents an increase 
in the refrigerating effect, the refrigerating 
effect per pound for the heat exchanger cycle is 

h c - h a . = 80.49 - 25.45 = 55.04 

The heat of compression per pound for the heat 
exchanger cycle is 

h a . - h d = 97.60 - 86.20 = 11.40 

Therefore, the coefficient of performance is 



K - h n . 55.04 



h,.-h, 11.40 



= 4.91 



The coefficient of performance of the satur- 
ated cycle is 4.88. Therefore, the coefficient of 
performance of the heat exchanger cycle is 
greater than that of the saturated cycle by only 



4.91 -4.88 



4.88 



x 100 =0.5% 



Depending upon the particular case, the 
coefficient of performance of a cycle employing 
a heat exchanger may be either greater than, 
less than, or the same as that of a saturated 
cycle operating between the same pressure 
limits. In any event, the difference is negligible, 
and it is evident that the advantages accruing 



116 PRINCIPLES OF REFRIGERATION 



20* vaporizing 
temperature 



Saturated 
suctio n f 
vapor- v 
20' 



Subcooled 
liquid-75"F~ 




Saturated 
liquid-lOOT 



Fig. 8-8. Flow diagram of refrigera- 
tion cycle illustrating the use of a 
liquid-suction heat exchanger. 



100* condensing 
temperature 

from the subcooling of the liquid in the heat 
exchanger are approximately offset by the 
disadvantages of superheating the vapor. 
Theoretically, then, the use of a heat exchanger 
cannot be justified on the basis of an increase in 
system capacity and efficiency. However, since 
in actual practice a refrigerating system does 
not (cannot) operate on a simple saturated 
cycle, this does not represent a true appraisal of 
the practical value of the heat exchanger. 
In an actual cycle, the suction vapor will 



131.16 



always become superheated before the com- 
pression process begins because nothing can be 
done to prevent it. This is true even if no 
superheating takes place either in the evapora- 
tor or in the suction line and the vapor reaches 
the inlet of the compressor at the vaporizing 
temperature. As the cold suction vapor flows 
into the compressor, it will become superheated 
by absorbing heat from the hot cylinder walls. 
Since the superheating in the compressor cyl- 
inder will occur before the compression process 



& 




35.75 , 



Enthalpy (Btu/lb) 

Hg. 8-». Ph diagrams comparing simple saturated cycle to cycle employing a liquid-suction heat exchanger. 
The amount of subcooling is equal to the amount of superheating. (Refrigerant- 1 2.) 



begins, the effect of the superheating on cycle 
efficiency will be approximately the same as if 
the superheating occurred in the suction line 
without producing useful cooling.* 

The disadvantages resulting from allowing the 
suction to become superheated without pro- 
ducing useful cooling have already been pointed 
out. Obviously, then, since superheating of the 
suction vapor is unavoidable in an actual cycle, 
whether or not a heat exchanger is used, any 
practical means of causing the vapor to become 
superheated in such a way that useful cooling 
results are worthwhile. Hence, the value of a 
heat exchanger lies in the fact that it provides a 
method of superheating the vapor so that useful 
cooling results. For this reason, the effect of a 
heat exchanger on cycle efficiency can be 
evaluated only by comparing the heat exchanger 
cycle to one in which the vapor is superheated 
without producing useful cooling. 

The maximum amount of heat exchange 
which can take place between the liquid and the 
vapor in the heat exchanger depends on the 
initial temperatures of the liquid and the vapor 
as they enter the heat exchanger and on the 
length of time they are in contact with each 
other. 

The greater the difference in temperature, the 
greater is the exchange of heat for any given 
period of contact. Thus, the lower the vaporiz- 
ing temperature and the higher the condensing 
temperature, the greater is the possible heat 
exchange. Theoretically, if the two fluids 
remained in contact for a sufficient length of 
time, they would leave the heat exchanger at the 
same temperature. In actual practice, this is 
not possible. However, the longer the two 
fluids stay in contact, the more nearly the two 
temperatures will approach one another. Since 
the specific heat of the vapor is less than that of 
the liquid, the rise in the temperature of the 

* It will be shown later that some advantages 
accrue from superheating which takes place in the 
compressor: (1) When the suction vapor absorbs 
heat from the cylinder walls, the cylinder wall tem- 
perature is lowered somewhat and this brings about 
a desirable change in the path of the compression 
process. However, the change is slight and is 
difficult to evaluate. (2) When hermetic motor- 
compressor assemblies are used, the suction vapor 
should reach the compressor at a relatively low 
temperature in order to help cool the motor windings. 



ACTUAL REFRIGERATING CYCLES 117 

vapor is always greater than the reduction in the 
temperature of the liquid. For instance, the 
specific heat of R-12 liquid is approximately 
0.24 Btu per pound, whereas the specific heat of 
the vapor is 0.15 Btu per pound. This means 
that the temperature reduction of the liquid will 
be approximately 62% (0.15/0.24) of the rise in 
the temperature of the vapor, or that for each 
24° F rise in the temperature of the vapor, the 
temperature of the liquid will be reduced 15° F. 

For the heat exchanger cycle in Fig. 8-9, the 
vapor absorbs 5.71 Btu per pound in super- 
heating from 20° F to 60° F. Assuming that all 
of the superheating takes place in the heat 
exchanger, the heat given up by the liquid is 5.71 
Btu, so that the temperature of the liquid is 
reduced 23.8° F (5.71/0.24) as the liquid passes 
through the heat exchanger. 
8-9. The Effect of Pressure Losses Resulting 
from Friction. In overcoming friction, both 
internal (within the fluid) and external (surface), 
the refrigerant experiences a drop in pressure 
while flowing through the piping, evaporator, 
condenser, receiver, and through the valves and 
passages of the compressor (Fig. 8-10). 

A Ph diagram of an actual cycle, illustrating 
the loss in pressure occurring in the various 
parts of the system, is shown in Fig. 8-11. To 
simplify the diagram, no superheating or sub- 
cooling is shown and a simple saturated cycle is 
drawn in for comparison. 

Line J9'-C'represents the vaporizing process in 
the evaporator during which the refrigerant 
undergoes a drop in pressure of 5.5 psi. Whereas 
the pressure and saturation temperature of the 
liquid-vapor mixture at the evaporator inlet is 
38.58 psia and 24° F, respectively, the pressure of 
the saturated vapor leaving the evaporator is 
33.08 psia, corresponding to a saturation tem- 
perature of 16° F. The average vaporizing tem- 
perature in the evaporator is 20° F, the same as 
that of the saturated cycle. 

As a result of the drop in pressure in the 
evaporator, the vapor leaves the evaporator at a 
lower pressure and saturation temperature and 
with a greater specific volume than if no drop in 
pressure occurred. 

The refrigerating effect per pound and the 
weight of refrigerant circulated per minute per 
ton are approximately the same for both cycles, 
but because of the greater specific volume the 
volume of vapor handled by the compressor per 



118 PRINCIPLES OF REFRIGERATION 



38.58 psia 



24* F (sat. temp.) * 

Pressure drop through 
evaporator, 5.5 psi 



Average evaporating 

temperature and 

pressure 
35.75 psia, 20° F 

33.08 psia// 
16* F (sat. temp.) 



c 



c 



£L 




, 158.9 psia 

1 114"F (sat. temp.) 
Pressure drop through 



discharge valves, 8.2 psi 



V 




90° F (sat. temp.) 



Pressure drop through 
liquid line, 77.3 psi 



150.7 psia 



|TTlO°F (sat. temp.) 

Pressure drop through 

hot gas line and 
condensers, 19.1 psi 

— *^ 

\ 131.6 psia 

100°F (sat. temp.) 



Average condensing 

temperature and pressure 

139 psia, 104° F 



Fig. 8-10. Flow diagram illustrating the effect of pressure drop in various parts of the system. Pressure drops 
are exaggerated for clarity. (Refrigerant- 1 2.) 



minute per ton is greater for the cycle experi- 
encing the pressure drop. Too, because of the 
lower pressure of the vapor leaving the evapora- 
tor, the vapor must be compressed through a 
greater pressure range during the compression 
process, so that the horsepower per ton is also 
greater for the cycle undergoing the drop in 
pressure. 

Line C'-C" represents the drop in pressure 
experienced by the suction vapor in flowing 



through the suction line from the evaporator to 
the compressor inlet. Like pressure drop in the 
evaporator, pressure drop in the suction line 
causes the suction vapor to reach the compressor 
at a lower pressure and in an expanded condition 
so that the volume of vapor compressed per 
minute per ton and the horsepower per ton are 
both increased. 

It is evident that the drop in pressure both in 
the evaporator and in the suction line should be 




Enthalpy (Btu/lb) 



Pressure drop 

1. Compressor discharge valves 4. Evaporator 

2. Discharge line and condenser 5. Suction line 

3. Liquid line 6. Compressor suction valves 



Fig. 8-11. Ph diagram of refrig- 
eration cycle illustrating the 
effect of pressure losses in the 
various parts of the system. 
A simple saturated cycle is 
drawn in for comparison. 
(Refrigerant- 1 2.) 



ACTUAL REFRIGERATING CYCLES 119 



kept to an absolute minimum in order to obtain 
the best possible cycle efficiency. This applies 
also to heat exchangers or any other auxiliary 
device intended for installation in the suction 
line. 

In Fig. 8-11, the pressure drops are exagger- 
ated for clarity. Ordinarily, good evaporator 
design limits the pressure drop across the 
evaporator to 2 or 3 psi. Ideally, the suction line 
should be designed so that the pressure drop is 
between 1 and 2 psi. 



against the spring-loading and to force the vapor 
out through the discharge valves and passages 
of the compressor into the discharge line. 

Line D'-A represents the drop in pressure 
resulting from the flow of the refrigerant through 
the discharge line and condenser. That part of 
line D'-A which represents the flow through the 
discharge line will vary with the particular case, 
since the discharge line may be either quite long 
or very short, depending upon the application. 
In any event, the result of the pressure drop will 



Fig. 8-12. Ph diagram of actual 
refrigeration cycle illustrating 
effects of subcooling, super- 
heating, and losses in pressure. 
A simple saturated cycle is 
drawn in for comparison. 
(Refrigerant- 1 2). 



I. 




Enthalpy (Btu/lb) 



Pressure drop 

1. Compressor discharge yalyes 4. Evaporator 

2. Discharge line end condenser 5. Suction line 

3. Liquid line 6. Compressor suction valves 



Line C-C " represents the drop in pressure that 
the suction vapor undergoes in flowing through 
the suction valves and passages of the com- 
pressor into the cylinder. The result of the drop 
in pressure through the valves and passages on 
the suction side of the compressor is the same as 
if the drop occurred in the suction line, and the 
effect on cycle efficiency is the same. Here again, 
good design requires that the drop in pressure be 
kept to a practical minimum. 

Line C-D " represents the compression process 
for the cycle undergoing the pressure drops. 
Notice that the vapor in the cylinder is com- 
pressed to a pressure considerably above the 
average condensing pressure. It is shown later 
that this is necessary in order to force the vapor 
out of the cylinder through the discharge valves 
against the condensing pressure and against the 
additional pressure occasioned by the spring- 
loading of the discharge valves. 

Line D"-D' represents the drop in pressure 
required to force the discharge valves open 



be the same. Any drop in pressure occurring on 
the discharge side of the compressor (in the dis- 
charge valves and passages, in the discharge line, 
and in the condenser) will have the effect of 
raising the discharge pressure and thereby 
increasing the work of compression and the 
horsepower per ton. 

Line A- A' represents the pressure drop result- 
ing from the flow of the refrigerant through the 
receiver tank and liquid line. Since the refrig- 
erant at A' is a saturated liquid, the temperature 
of the liquid must decrease as the pressure 
decreases. If the liquid is not subcooled by 
giving up heat to an external sink as its pressure 
drops, a portion of the liquid must flash into a 
vapor in the liquid line in order to provide the 
required cooling of the liquid. Notice that point 
A" lies in the region of phase-change, indicating 
that a portion of the refrigerant is a vapor at 
this point. 

Despite the flashing of the liquid and the drop 
in temperature coincident with the drop in 



120 PRINCIPLES OF REFRIGERATION 



pressure in the liquid line, the drop in pressure 
in the liquid line has no effect on cycle efficiency. 
The pressure and temperature of die liquid must 
be reduced to the vaporizing condition before it 
enters the evaporator in any case. The fact that 
a part of this takes place in the liquid line rather 
than in the refrigerant control has no direct 
effect on the efficiency of the system. It does, 
however, reduce the capacity of both the liquid 
line and the refrigerant control. Furthermore, 
passage of vapor through the refrigerant control 
will eventually cause damage to the refrigerant 
control by eroding the valve needle and seat. 

Ordinarily, even without the use of a heat 
exchanger, sufficient subcooling of the liquid 
will occur in the liquid line to prevent the 
flashing of the liquid if the drop in pressure in 
the line is not excessive. Flashing of the liquid 
in the liquid line will usually not take place when 
the drop in the line does not exceed 5 psi. 

The effect of pressure drop in the lines and in 
the other parts of the system is discussed more 
fully later in the appropriate chapters. 

A Ph diagram of a typical refrigeratioircycle, 
which illustrated the combined effects of pressure 
drop, subcooling, and superheating, is compared 
to the PA diagram of the simple saturated cycle 
in Fig. 8-12. 

PROBLEMS 

1. The vaporizing and condensing temperature 
of a Refrigerant-12 system are 40° F and 1 10° F, 
respectively. The suction vapor is superheated 



to 70° F in the suction line, whereas the liquid 
is subcooled to 90° F by giving off heat to the 
ambient air. Determine: 

(a) The refrigerating effect per pound. 

Ans. 54.01 Btu/lb 

(b) The weight of refrigerant circulated per 
minute per ton. Ans. 3.70 lb/min/ton 

(c) The volume of vapor compressed per 
minute per ton. Ans. 2.93 cu ft/min/ton 

(d) The loss of refrigerating effect per pound 
in the refrigerant control. 

Ans. 11.7 Btu/lb 

(c) The quantity of superheat in the suction 

vapor. Ans. 4.39 Btu/lb 

(/) The gain in refrigerating effect per pound 

resulting from the liquid subcooling. 

Ans. 4.93 Btu/lb 
(g ) The adiabatic discharge temperature. 

Ans. 138.5° F 
(A) The heat of compression per pound. 

Ans. 9 Btu/lb 

(0 The heat of compression per minute per 

ton. Ans. 33.3 Btu/min/ton 

(J) The work of compression per minute per 

ton. Ans. 25.907 lb/min/ton 

(At) The theoretical horsepower per ton. 

Ans. 0.755 hp/ton 

(/)The heat rejected at the condenser per 

pound. Arts. 67 .4 Btu/lb 

(m) The heat rejected at the condenser per ton. 

Ans. 249.38 Btu/min/ton 

(») The coefficient of performance. Ans. 6 

Note: Some of the properties of the refriger- 
ant at various points in the cycle must be 
determined from the Ph chart in Fig. 7-1. 



9 

Survey 

of Refrigeration 

Applications 



9-1. History and Scope of the Industry. In 

the early days of mechanical refrigeration, the 
equipment available was bulky, expensive, and 
not too efficient. Also it was of such a nature as 
to require that a mechanic or operating engineer 
be on duty at all times. This limited the use of 
mechanical refrigeration to a few large applica- 
tions such as ice plants, meat packing plants, and 
large storage warehouses. 

In the span of only a few decades refrigeration 
has grown into the giant and rapidly expanding 
industry that it is today. This explosive growth 
came about as the result of several factors. First, 
with the development of precision manufacturing 
methods, it became possible to produce smaller, 
more efficient equipment. This, along with the 
development of "safe" refrigerants and the 
invention of the fractional horsepower electric 
motor, made possible the small refrigerating 
unit which is so widely used at the present time 
in such applications as domestic refrigerators 
and freezers, small air conditioners, and com- 
mercial fixtures. Today, there are few homes or 
business establishments in the United States that 
cannot boast of one or more mechanical 
refrigeration units of some sort. 

Few people outside of those directly connected 
with the industry are aware of the significant 
part that refrigeration has played in the develop- 
ment of the highly technical society that 



America is today, nor do they realize the extent 
to which such a society is dependent upon 
mechanical refrigeration for its very existence. 
It would not be possible, for instance, to pre- 
serve food in sufficient quantities to feed the 
growing urban population without mechanical 
refrigeration. Too, many of the large buildings 
which house much of the nation's business and 
industry would become untenable in the summer 
months because of the heat if they were not air 
conditioned with mechanical refrigerating 
equipment. 

In addition to the better known applications 
of refrigeration, such as comfort air conditioning 
and the processing, freezing, storage, transpor- 
tation, and display of perishable products, 
mechanical refrigeration is used in the processing 
or manufacturing of almost every article or 
commodity on the market today. The list of 
processes or products made possible or improved 
through the use of mechanical refrigeration is 
almost endless. For example, refrigeration has 
made possible the building of huge dams which 
are vital to large-scale reclamation and hydro- 
electric projects. It has made possible the con- 
struction of roads and tunnels and the sinking of 
foundation and mining shafts through and across 
unstable ground formations. It has made 
possible the production of plastics, synthetic 
rubber, and many other new and useful mate- 
rials and products. Because of mechanical 
refrigeration, bakers can get more loaves of 
bread from a barrel of flour, textile and paper 
manufacturers can speed up their machines and 
get more production, and better methods of 
hardening steels for machine tools are available. 
These represent only a few of the hundreds of 
ways in which mechanical refrigeration is now 
being used and many new uses are being found 
each year. In fact, the only thing slowing the 
growth of the refrigeration industry at the 
present time is the lack of an adequate supply of 
trained technical manpower. 
9-2. Classification of Applications. For con- 
venience of study, refrigeration applications may 
be grouped into six general categories: (1) 
domestic refrigeration, (2) commercial refrigera- 
tion, (3) industrial refrigeration, (4) marine and 
transportation refrigeration, (5) comfort air con- 
ditioning, and (6) industrial air conditioning. It 
will be apparent in the discussion which follows 
that the exact limits of these areas are not 



121 



122 PRINCIPLES OF REFRIGERATION 

precisely defined and that there is considerable 
overlapping between the several areas. 
9-3. Domestic Refrigeration. Domestic re- 
frigeration is rather limited in scope, being 
concerned primarily with household refrigera- 
tors and home freezers. However, because the 
number of units in service is quite large, 
domestic refrigeration represents a significant 
portion of the refrigeration industry. 

Domesticunits are usually small in size, having 
horsepower ratings of between ^ and J hp, and 
are of the hermetically sealed type. Since these 
applications are familiar to everyone, they will 
not be described further here. However, the 
problems encountered in the design and main- 
tenance of these units are discussed in appro- 
priate places in the chapters which follow. 
9-4. Com.nercial Refrigeration. Commer- 
cial refrigeration is concerned with the designing, 
installation, and maintenance of refrigerated 
fixtures of the type used by retail stores, res- 
taurants, hotels, and institutions for the storing, 
displaying, processing, and dispensing of perish- 
able commodities of all types. Commercial 
refrigeration fixtures are described in more 
detail later in this chapter. 
9-5. Industrial Refrigeration. Industrial 
refrigeration is often confused with commercial 
refrigeration because the division between these 
two areas is not clearly defined. As a general 
rule, industrial applications are larger in size 
than commercial applications and have the dis- 
tinguishing feature of requiring an attendant on 
duty, usually a licensed operating engineer. 
Typical industrial applications are ice plants, 
large food-packing plants (meat, fish, poultry, 
frozen foods, etc.), breweries, creameries, and 
industrial plants, such as oil refineries, chemical 
plants, rubber plants, etc. Industrial refrigera- 
tion includes also those applications concerned 
with the construction industry as described in 
Section 9-1. 

9-6. Marine and Transportation Refrigera- 
tion. Applications falling into this category 
could be listed partly under commercial refrig- 
eration and partly under industrial refrigera- 
tion. However, both these areas of specialization 
have grown to sufficient size to warrant special 
mention. 

Marine refrigeration, of course, refers to 
refrigeration aboard marine vessels and includes, 
for example, refrigeration for fishing boats and 



for vessels transporting perishable cargo as well 
as refrigeration for the ship's stores on vessels of 

all kinds. 

Transportation refrigeration is concerned 
with refrigeration equipment as it is applied to 
trucks, both long distance transports and local 
delivery, and to refrigerated railway cars. Typi- 
cal refrigerated truck bodies are shown in Fig. 

11-8. 

9-7. Air Conditioning. As the name implies, 
air conditioning is concerned with the condition 
of the air in some designated area or space. This 
usually involves control not only of the space 
temperature but also of space humidity and air 
motion, along with the filtering and cleaning of 
the air. 

Air conditioning applications are of two types, 
either comfort or industrial, according to their 
purpose. Any air conditioning which has as its 
primary function the conditioning of air for 
human comfort is called comfort air condition- 
ing. Typical installations of comfort air 
conditioning are in homes, schools, offices, 
churches, hotels, retail stores, public buildings, 
factories, automobiles, buses, trains, planes, 
ships, etc. 

On the other hand, any air conditioning which 
does not have as its primary purpose the con- 
ditioning of air for human comfort is called 
industrial air conditioning. This does not 
necessarily mean that industrial air conditioning 
systems cannot serve as comfort air conditioning 
coincidentally with their primary function. 
Often this is the case, although not always so. 
The applications of industrial air conditioning 
are almost without limit both in number and in 
variety. Generally speaking, the functions of 
industrial air conditioning are to; (1) control 
the moisture content of hydroscopic materials; 
(2) govern the rate of chemical and biochemical 
reactions; (3) limit the variations in the size of 
precision manufactured articles because of ther- 
mal expansion and contraction ; and (4) provide 
clean, filtered air which is often essential to 
trouble-free operation and to the production of 
quality products. 

9-8. Food Preservation. The preservation of 
perishable commodities, particularly foodstuffs, 
is one of the most common uses of mechanical 
refrigeration. As such, it is a subject which 
should be given consideration in any compre- 
hensive study of refrigeration. 



SURVEY OF REFRIGERATION APPLICATIONS 123 



At the present time, food preservation is more 
important than ever before in man's history. 
Today's large urban populations require tremen- 
dous quantities of food, which for the most part 
must be produced and processed in outlying 
areas. Naturally, these foodstuffs must be kept 
in a preserved condition during transit and sub- 
sequent storage until they are finally consumed. 
This may be a matter of hours, days, weeks, 
months, or even years in some cases. Too, many 
products, particularly fruit and vegetables, are 
seasonal. Since they are produced only during 
certain seasons of the year, they must be stored 
and preserved if they are to be made available 
the year round. 

As a matter of life or death, the preservation 
of food has long been one of man's most pressing 
problems. Almost from the very beginning of 
man's existence on earth, it became necessary 
for him to find ways of preserving food during 
seasons of abundance in order to live through 
seasons of scarcity. It is only natural, then, that 
man should discover and develop such methods 
of food preservation as drying, smoking, pick- 
ling, and salting long before he had any know- 
ledge of the causes of food spoilage. These 
rather primitive methods are still widely used 
today, not only in backward societies where no 
other means are available but also in the most 
modern societies where they serve to supplement 
the more modern methods of food preservation. 
For instance, millions of pounds of dehydrated 
(dried) fruit, milk, eggs, fish, meat, potatoes, etc., 
are consumed in the United States each year, 
along with huge quantities of smoked, pickled, 
and salted products, such as ham, bacon, and 
sausage, to name only a few. However, although 
these older methods are entirely adequate for the 
preservation of certain types of food, and often 
produce very unusual and tasty products which 
would not otherwise be available, they nonethe- 
less have inherent disadvantages which limit 
their usefulness. Since by their very nature they 
bring about severe changes in appearance, taste, 
and odor, which in many cases are objectionable, 
they are not universally adaptable for the pres- 
ervation of all types of food products. Further- 
more, the keeping qualities of food preserved by 
such methods are definitely limited as to time. 
Therefore, where a product is to be preserved 
indefinitely or for a long period of time, some 
other means of preservation must be utilized. 



The invention of the microscope and the sub- 
sequent discovery of microorganisms as a major 
cause of food spoilage led to the development of 
canning in France during the time of Napoleon. 
With the invention of canning, man found a way 
to preserve food of all kinds in large quantities 
and for indefinite periods of time. Canned foods 
have the advantage of being entirely imperish- 
able, easily processed, and convenient to handle 
and store. Today, more food is preserved by 
canning than by all other methods combined. 
The one big disadvantage of canning is that 
canned foods must be heat-sterilized, which 
frequently results in overcooking. Hence, al- 
though canned foods often have a distinctive and 
delicious flavor all their own, they usually differ 
greatly from the original fresh product. 

The only means of preserving food in its 
original fresh state is by refrigeration. This, of 
course, is the principal advantage that refrigera- 
tion has over other methods of food preser- 
vation. However, refrigeration too has its 
disadvantages. For instance, when food is to 
be preserved by refrigeration, the refrigerating 
process must begin very soon after harvesting 
or killing and must be continuous until the food 
is finally consumed. Since this requires rela- 
tively expensive and bulky equipment, it is often 
both inconvenient and uneconomical. 

Obviously, then, there is no one method of 
food preservation which is best in all cases and 
the particular method used ih any one case will 
depend upon a number of factors, such as the 
type of product, the length of time the product 
is to be preserved, the purpose for which the 
product is to be used, the availability of trans- 
portation and storage equipment, etc. Very 
often it is necessary to employ several methods 
simultaneously in order to obtain the desired 
results. 

9-9. Deterioration and Spoilage. Since the 
preservation of food is simply a matter of pre- 
venting or retarding deterioration and spoilage 
regardless of the method used, a good knowledge 
of the causes of deterioration and spoilage is 
a prerequisite to the study of preservation 
methods. 

It should be recognized at the outset that 
there are degrees of quality and that all perish- 
able foods pass through various stages of 
deterioration before becoming unfit for con- 
sumption. In most cases, the objective in the 



124 PRINCIPLES OF REFRIGERATION 

preservation of food is not only to preserve the 
foodstuff in an edible condition but also to 
preserve it as nearly as possible at the peak of its 
quality with respect to appearance, odor, taste, 
and vitamin content. Except for a few processed 
foods, this usually means maintaining the food- 
stuff as nearly as possible in its original fresh 
state. 

Any deterioration sufficient to cause a detect- 
able change in the appearance, odor, or taste of 
fresh foods immediately reduces the commercial 
value of the product and thereby represents an 
economic loss. Consider, for example, wilted 
vegetables or overripe fruit. Although their 
edibility is little impaired, an undesirable change 
in their appearance has been brought about 
which usually requires that they be disposed of 
at a reduced price. Too, since they are well on 
their way to eventual spoilage, their keeping 
qualities are greatly reduced and they must be 
consumed or processed immediately or become 
a total loss. 

For obvious reasons, maintaining the vitamin 
content at the highest possible level is always an 
important factor in the processing and/or pres- 
ervation of all food products. In fact, many 
food processors, such as bakers and dairymen, 
are now adding vitamins to their product to 
replace those which are lost during processing. 
Fresh vegetables, fruit, and fruit juices are some 
of the food products which suffer heavy losses in 
vitamin content very quickly if they are not 
handled and protected properly. Although the 
loss of vitamin content is not something which 
in itself is apparent, in many fresh foods it is 
usually accompanied by recognizable changes in 
appearance, odor, or taste, such as, for instance, 
wilting in leafy, green vegetables. 

For the most part, the deterioration and even- 
tual spoilage of perishable food are caused by a 
series of complex chemical changes which take 
place in the foodstuff after harvesting or killing. 
These chemical changes are brought about by 
both internal and external agents. The former 
are the natural enzymes which are inherent in all 
organic materials, whereas the latter are micro- 
organisms which grow in and on the surface of 
the foodstuff. Although either agent alone is 
capable of bringing about the total destruction 
of a food product, both agents are involved in 
most cases of food spoilage. In any event, the 
activity of both of these spoilage agents must be 



either eliminated or effectively controlled if the 
foodstuff is to be adequately preserved. 
9-10. Enzymes. Enzymes are complex, pro- 
tein-like, chemical substances. Not yet fully 
understood, they are probably best described 
as chemical catalytic agents which are capable 
of bringing about chemical changes in organic 
materials. There are many different kinds of 
enzymes and each one is specialized in that it 
produces only one specific chemical' reaction. 
In general, enzymes are identified either by the 
substance upon which they act or by the result 
of their action. For instance, the enzyme, 
lactase, is so known because it acts to convert 
lactose (milk sugar) to lactic acid. This par- 
ticular process is called lactic acid fermentation 
and is the one principally responsible for the 
"souring" of milk. Enzymes associated with 
the various types of fermentation are sometimes 
called ferments. 

Essential in the chemistry of all living pro- 
cesses, enzymes are normally present in all 
organic materials (the cell tissue of all plants and 
animals, both living and dead). They are manu- 
factured by all living cells to help carry on the 
various living activities of the cell, such as res- 
piration, digestion, growth, and reproduction, 
and they play an important part in such things 
as the sprouting of seeds, the growth of plants 
and animals, the ripening of fruit, and the 
digestive processes of animals, including man. 
However, enzymes are catabolic as well as ana- 
bolic. That is, they act to destroy dead cell tissue 
as well as to maintain live cell tissue. In fact, 
enzymes are the agents primarily responsible 
for the decay and decomposition of all organic 
materials, as, for example, the putrification of 
meat and fish and the rotting of fruit and 
vegetables. 

Whether their action is catabolic or anabolic, 
enzymes are nearly always destructive to perish- 
able foods. Therefore, except in those few 
special cases where fermentation or putrification 
is a part of the processing, enzymic action must 
be either eliminated entirely or severely inhibited 
if the product is to be preserved in good con- 
dition. Fortunately, enzymes are sensitive to 
the conditions of the surrounding media, par- 
ticularly with regard to the temperature and the 
degree of acidity or alkalinity, which provides a 
practical means of controlling enzymic activity. 
Enzymes are completely destroyed by high 



SURVEY OF REFRIGERATION APPLICATIONS 125 



temperatures that alter the composition of the 
organic material in which they exist. Since most 
enzymes are eliminated at temperatures above 
160° F, cooking a food substance completely 
destroys the enzymes contained therein. On the 
other hand, enzymes are very resistant to low 
temperatures and their activity may continue at 
a slow rate even at temperatures below 0° F. 
However, it is a well-known fact that the rate of 
chemical reaction decreases as the temperature 
decreases. Hence, although the enzymes are not 
destroyed, their activity is greatly reduced at low 
temperatures, particularly temperatures below 
the freezing point of water. 

Enzymic action is greatest in the presence of 
free oxygen (as in the air) and decreases as the 
oxygen supply diminishes. 

With regard to the degree of acidity or alka- 
linity, some enzymes require acid surroundings, 
whereas others prefer neutral or alkaline en- 
vironments. Those requiring acidity are de- 
stroyed by alkalinity and those requiring 
alkalinity are likewise destroyed by acidity. 

Although an organic substance can be com- 
pletely destroyed and decomposed solely by the 
action of its own natural enzymes, a process 
known as autolysis (self-destruction), this sel- 
dom occurs. More often, the natural enzymes 
are aided in their destructive action by enzymes 
secreted by microorganisms. 
9-11. Microorganisms. The term micro- 
organism is used to cover a whole group of 
minute plants and animals of microscopic and 
submicroscopic size, of which only the following 
three are of particular interest in the study of 
food preservation: (1) bacteria, (2) yeasts, and 
(3) molds. These tiny organisms are found in 
large numbers everywhere — in the air, in the 
ground, in water, in and on the bodies of plants 
and animals, and in every other place where con- 
ditions are such that living organisms can survive. 

Because they secrete enzymes which attack 
the organic materials upon which they grow, 
microorganisms are agents of fermentation, 
purification, and decay. As such, they are both 
beneficial and harmful to mankind. Then- 
growth in and on the surface of perishable foods 
causes complex chemical changes in the food 
substance which usually results in undesirable 
alterations in the taste, odor, and appearance of 
the food and which, if allowed to continue for 
any length of time, will render the food unfit 



for consumption. Too, some microorganisms 
secrete poisonous substances (toxins) which are 
extremely dangerous to health, causing poison- 
ing, disease, and often death. 

On the other hand, microorganisms have 
many useful and necessary functions. As a 
matter of fact, if it were not for the work of 
microorganisms, life of any kind would not be 
possible. Since decay and decomposition of all 
dead animal tissue are essential to make space 
available for new life and growth, the decaying 
action of microorganisms is indispensable to the 
life cycle. 

Of all living things, only green plants (those 
containing chlorophyll) are capable of using 
inorganic materials as food for building their 
cell tissue. Through a process called photo- 
synthesis, green plants are able to utilize the 
radiant energy of the sun to combine carbon 
dioxide from the air with water and mineral 
salts from the soil and thereby manufacture 
from inorganic materials the organic compounds 
which make up their cell tissue. 

Conversely, all animals and all plants without 
chlorophyll (fungi) require organic materials 
(those containing carbon) for food to carry on 
their life activities. Consequently, they must of 
necessity feed upon the cell tissue of other plants 
and animals (either living or dead) and are, 
therefore, dependent either directly or indirectly 
on green plants as a source of the organic 
materials they need for life and growth. 

It is evident, then, that should the supply of 
inorganic materials in the soil, which serve as 
food for green plants, ever become exhausted, 
all life would soon disappear from the earth. 
This is not likely to happen, however, since 
microorganisms, as a part of their own living 
process, are continuously replenishing the supply 
of inorganic materials in the soil. 

With the exception of a few types of soil 
bacteria, all microorganisms need organic 
materials as food to carry on the living process. 
In most cases, they obtain these materials by 
decomposing animal wastes and the tissue of 
dead animals and plants. In the process of 
decomposition, the complex organic compounds 
which make up the tissue of animals and plants 
are broken down step by step and are eventually 
reduced to simple inorganic materials which are 
returned to the soil to be used as food by the 
green plants. 



126 PRINCIPLES OF REFRIGERATION 

In addition to the important part they play in 
the "food chain" by helping to keep essential 
materials in circulation, microorganisms ale 
necessary in the processing of certain fermented 
foods and other commodities. For example, 
bacteria are responsible for the lactic acid fer- 
mentation required in the processing of pickles, 
olives, cocoa, coffee, sauerkraut, ensilage, and 
certain sour milk products, such as butter, 
cheese, buttermilk, yogurt, etc., and for the 
acetic acid fermentation necessary in the pro- 
duction of vinegar from various alcohols. 
Bacterial action is useful also in the processing 
of certain other commodities such as leather, 
linen, hemp, and tobacco, and in the treatment 
of industrial wastes of organic composition. 

Yeasts, because of their ability to produce 
alcoholic fermentation, are of immeasurable 
value to the brewing and wine-making industries 
and to the production of alcohols of all kinds. 
Too, everyone is aware of the importance of 
yeast in the baking industry. 

The chief commercial uses of molds are in the 
processing of certain types of cheeses and, more 
important, in the production of antibiotics, such 
as penicillin and aureomycin. 

Despite their many useful and necessary 
functions, the fact remains that microorganisms 
are destructive to perishable foods. Hence, 
their activity, like that of the natural enzymes, 
must be effectively controlled if deterioration and 
spoilage of the food substance are to be avoided. 

Since each type of microorganism differs 
somewhat in both nature and behavior, it is 
worthwhile to examine each type separately. 
9-12. Bacteria. Bacteria are a very simple 
form of plant life, being made up of one single 
living cell. Reproduction is accomplished by cell 
division. On reaching maturity, the bacterium 
divides into two separate and equal cells, each 
of which in turn grows to maturity and divides 
into two cells. Bacteria grow and reproduce at 
an enormous rate. Under ideal conditions, a 
bacterium can grow into maturity and repro- 
duce in as little as 20 to 30 min. At this rate a 
single bacterium is capable of producing as 
many as 34,000,000,000,000 descendants in a 
24-hr period. Fortunately, however, the life 
cycle of bacteria is relatively short, being a 
matter of minutes or hours, so that even under 
ideal conditions they cannot multiply at any- 
where near this rate. 



The rate at which bacteria and other micro- 
organisms grow and reproduce depends upon 
such environmental conditions as temperature, 
light, and the degree of acidity or alkalinity, and 
upon the availability of oxygen, moisture, and 
an adequate supply of soluble food. However, 
there are many species of bacteria and they 
differ greatly both in their choice of environment 
and in the effect they have on their environment. 
Like the higher forms of plant life, all species of 
bacteria are not equally hardy with respect to 
surviving adverse conditions of environment, 
nor do all species thrive equally well under the 
same environmental conditions. Some species 
prefer conditions which are entirely fatal to 
others. Too, some bacteria are spore-formers. 
The spore is formed within the bacteria cell and 
is protected by a heavy covering or wall. In the 
spore state, which is actually a resting or 
dormant phase of the organism, bacteria are 
extremely resistant to unfavorable conditions of 
environment and can survive in this state almost 
indefinitely. The spore will usually germinate 
whenever conditions become favorable for the 
organism to carry on its living activities. 

Most bacteria are saprophytes. That is, they 
are "free living" and feed only on animal wastes 
and on the dead tissue of animals and plants. 
Some, however, are parasites and require a 
living host. Most pathogenic bacteria (those 
causing infection and disease) are of the para- 
sitic type. In the absence of a living host, some 
parasitic bacteria can live as saprophytes. Like- 
wise, some saprophytes can live as parasites 
when the need arises. 

Since bacteria are not able to digest insoluble 
food substances, they require food in a soluble 
form. For this reason, most bacteria secrete 
enzymes which are capable of rendering in- 
soluble compounds into a soluble state, thereby 
making these materials available to the bacteria 
as food. The deterioration of perishable foods 
by bacteria growth is a direct result of the action 
of these bacterial enzymes. 

Bacteria, like all other living things, require 
moisture as well as food to carry on their life 
activities. As in other things, bacteria vary 
considerably in their ability to resist drought. 
Although most species are readily destroyed by 
drying and will succumb within a few hours, the 
more hardy species are able to resist drought 
for several days. Bacterial spores can withstand 



SURVEY OF REFRIGERATION APPLICATIONS 
The Growth of Bacteria in Milk in Various Periods 



127 







Time, hours 




Temp., °F 


24 


48 


96 


168 


32 


2,400 


2,100 


1,850 


1,400 


39 


2,500 


3,600 


218,000 


4,200,000 


46 


3,100 


12,000 


1,480,000 




50 


11,600 


540,000 






60 


180,000 


28,000,000 






86 


1,400,000,000 









Fig. 9-1. From ASRE Data Book, Applications Volume, 1956-57. Reproduced by permission of the American 
Society of Heating, Refrigerating, and Air-Conditioning Engineers. 



drought almost indefinitely, but will remain 
dormant in the absence of moisture. 

In their need for oxygen, bacteria fall into two 
groups : (1) those which require free oxygen (air) 
and (2) those which can exist without free oxy- 
gen. Some species, although having a prefer- 
ence for one condition or the other, can live 
in the presence of free oxygen or in the absence 
of it. Those bacteria living without free oxygen 
obtain the needed oxygen through chemical 
reaction which reduces one compound while 
oxidizing another. Decomposition which occurs 
in the presence of free oxygen is known as decay, 
whereas decomposition which takes place in the 
absence of free oxygen is called putrification. 
One of the products of putrification is hydrogen 
sulfide, a foul-smelling gas which is frequently 
noted arising from decomposing animal 
carcasses. 

Bacteria are very sensitive to acidity or alka- 
linity and cannot survive in an either highly acid 
or highly alkaline environment. Most bacteria 
require either neutral or slightly alkaline sur- 
roundings, although some species prefer slightly 
acid conditions. Because bacteria prefer neutral 
or slightly alkaline surroundings, nonacid 
vegetables are especially subject to bacterial 
attack. 

Light, particularly direct sunlight, is harmful 
to all bacteria. Whereas visible light only 
inhibits their growth, ultraviolet light is actually 
fatal to bacteria. Since light rays, ultraviolet or 
otherwise, have no power of penetration, they 
are effective only in controlling surface bacteria. 
However, ultraviolet radiation (usually from 
direct sunlight), when combined with drying, 
provides an excellent means of controlling 
bacteria growth. 



For each species of bacteria there is an opti- 
mum temperature at which the bacteria will 
grow at the highest rate. Too, for each species 
there is a maximum and a minimum tempera- 
ture which will permit growth. At temperatures 
above the maximum, the bacteria are destroyed. 
At temperatures below the minimum, the bac- 
teria are rendered inactive or dormant. The 
optimum temperature for most saprophytes is 
usually between 75° F and 85° F, whereas the 
optimum temperature for parasites is around 
99° F or 100° F. A few species grow best at 
temperatures near the boiling point of water, 
whereas a few other types thrive best at tempera- 
tures near the freezing point. However, most 
species are either killed off or severely inhibited 
at these temperatures. The effect of temperature 
on the growth rate of bacteria is illustrated by 
the chart in Fig. 9-1 which shows the growth 
rate of bacteria in milk at various temperatures. 
In general, the growth rate of bacteria is con- 
siderably reduced by lowering the temperature. 
9-13. Yeasts. Yeasts are simple, one-cell plants 
of the fungus family. Of microscopic size, 
yeast cells are somewhat larger and more com- 
plex than the bacteria cells. Although a few 
yeasts reproduce by fission or by sexual process, 
reproduction is usually by budding. Starting as 
a small protrusion of the mature cell, the bud 
enlarges and finally separates from the mother 
cell. Under ideal conditions, budding is fre- 
quently so rapid that new buds are formed before 
separation occurs so that yeast clusters are 
formed. 

Like bacteria, yeasts are agents of fermenta- 
tion and decay. They secrete enzymes that bring 
about chemical changes in the food upon which 
they grow. Yeasts are noted for their ability 



128 PRINCIPLES OF REFRIGERATION 

to transform sugars into alcohol and carbon 
dioxide. Although destructive to fresh foods, 
particularly fruits and berries and their juices, 
the alcoholic fermentation produced by yeasts 
is essential to the baking, brewing, and wine- 
making industries. 

Yeasts are spore-formers, with as many as 
eight spores being formed within a single yeast 
cell. Yeasts are widespread in nature and yeast 
spores are invariably found in the air and on the 
skin of fruit and berries, for which they have a 
particular affinity. They usually spend the 
winter in the soil and are carried to the new fruit 
in the spring by insects or by the wind. 

Like bacteria, yeasts require air, food, and 
moisture for growth, and are sensitive to tem- 
perature and the degree of acidity or alkalinity 
in the environment. For the most part, yeasts 
prefer moderate temperatures and slight acidity. 
In general, yeasts are not as resistant to unfavor- 
able conditions as are bacteria, although they 
can grow in acid surroundings which inhibit 
most bacteria. Yeast spores, like those of bac- 
teria, are extremely hardy and can survive for 
long periods under adverse conditions. 
9-14. Molds. Molds, like yeasts, are simple 
plants of the fungi family. However, they are 
much more complex in structure than either 
bacteria or yeasts. Whereas the individual 
bacteria or yeast plants consist of one single 
cell, an individual mold plant is made up of a 
number of cells which are positioned end to end 
to form long, branching, threadlike fibers called 
hypha. The network which is formed by a mass 
of these threadlike fibers is called the mycelium 
and is easily visible to the naked eye. The 
hyphae of the mold plant are of two general 
types. Some are vegetative fibers which grow 
under the surface and act as roots to gather food 
for the plant, whereas others, called aerial 
hyphae, grow on the surface and produce the 
fruiting bodies. 

Molds reproduce by spore formation. The 
spores develop in three different ways, depending 
on the type of mold: (1) as round clusters 
within the fibrous hyphae network, (2) as a mass 
enclosed in a sac and attached to the end of 
aerial hyphae, and (3) as chainlike clusters on 
the end of aerial hyphae. In any case, a single 
mold plant is capable of producing thousands of 
spores which break free from the mother plant 
and float away with the slightest air motion. 



Mold spores are actually seeds and, under the 
proper conditions, will germinate and produce 
mold growth on any food substance with which 
they come in contact. Since they are carried 
about by air currents, mold spores are found 
almost everywhere and are particularly abundant 
in the air. 

Although molds are less resistant to high 
temperatures than are bacteria, they are more 
tolerant to low temperatures, growing freely at 
temperatures close to the freezing point of water. 
Mold growth is inhibited by temperatures below 
32° F, more from the lack of free moisture than 
from the effect of low temperature. All mold 
growth ceases at temperatures of 10° F and 
below. 

Molds flourish in dark, damp surroundings, 
particularly in still air. An abundant supply of 
oxygen is essential to mold growth, although a 
very few species can grow in the absence of 
oxygen. Conditions inside cold-storage rooms 
are often ideal for mold growth, especially in 
the wintertime. This problem can be overcome 
somewhat by maintaining good air circulation 
in the storage room, by the use of germicidal 
paints, and ultraviolet radiation, and by frequent 
scrubbing. 

Unlike bacteria, molds can thrive on foods 
containing relatively large amounts of sugars or 
acids. They are often found growing on acid 
fruits and on the surface of pickling vats, and 
are the most common cause of spoilage in 
apples and citrus fruits. 

9-15. Control of Spoilage Agents. Despite 
complications arising from the differences in the 
reaction of the various types of spoilage agents 
to specific conditions in the environment, con- 
trolling these conditions provides the only means 
of controlling these spoilage agents. Thus, all 
methods of food preservation must of necessity 
involve manipulation of the environment in and 
around the preserved product in order to pro- 
duce one or more conditions unfavorable to the 
continued activity of the spoilage agents. When 
the product is to be preserved for any length of 
time, the unfavorable conditions produced must 
be of sufficient severity to eliminate the spoilage 
agents entirely or at least render them ineffective 
or dormant. 

All types of spoilage agents are destroyed when 
subjected to high temperatures over a period of 
time. This principle is used in the preservation 



SURVEY OF REFRIGERATION APPLICATIONS 129 



of food by canning. The temperature of the 
product is raised to a level fatal to all spoilage 
agents and is maintained at this level until they 
are all destroyed. The product is then sealed 
in sterilized, air-tight containers to prevent 
recontamination. A product so processed will 
remain in a preserved state indefinitely. 

The exposure time required for the destruc- 
tion of all spoilage agents depends upon the 
temperature level. The higher the temperature 
level, the shorter is the exposure period required. 
In this regard, moist heat is more effective than 
dry heat because of its greater penetrating 
powers. When moist heat is used, the tempera- 
ture level required is lower and the processing 
period is shorter. Enzymes and all living micro- 
organisms are destroyed when exposed to the 
temperature of boiling water for approximately 
five minutes, but the more resistant bacteria 
spores may survive at this condition for several 
hours before succumbing. For this reason, some 
food products, particularly meats and nonacid 
vegetables, require long processing periods 
which frequently result in overcooking of the 
product. These products are usually processed 
under pressure so that the processing tempera- 
ture is increased and the processing time 
shortened. 

Another method of curtailing the activity of 
spoilage agents is to deprive them of the mois- 
ture and/or food which is necessary for then- 
continued activity. Both enzymes and micro- 
organisms require moisture to carry on their 
activities. Hence, removal of the free moisture 
from a product will severely limit their activities. 
The process of moisture removal is called drying 
(dehydration) and is one of the oldest methods of 
preserving foods. Drying is accomplished either 
naturally in the sun and air or artificially in 
ovens. Dried products which are stored in a 
cool, dry place will remain in good condition 
for long periods. 

Pickling is essentially a fermentation process, 
the end result of which is the exhaustion of the 
substances which serve as food for yeasts and 
bacteria. The product to be preserved by pick- 
ling is immersed in a salt brine solution and 
fermentation is allowed to take place, during 
which the sugars contained in the food product 
are converted to lactic acid, primarily through 
the action of lactic acid bacteria. 
Smoked products are preserved partially by 



the drying effect of the smoke and partially 
by antiseptics (primarily creosote) which are 
absorbed from the smoke. 

Too, some products are "cured" with sugar 
or salt which act as preservatives in that they 
create conditions unfavorable to the activity of 
spoilage agents. Some other frequently used 
preservatives are vinegar, borax, saltpeter, bon- 
zoate of soda, and various spices. A few of the 
products preserved in this manner are sugar- 
cured hams, salt pork, spiced fruits, certain 
beverages, jellies, jams, and preserves. 
9-16. Preservation by Refrigeration. The 
preservation of perishables by refrigeration in- 
volves the use of low temperature as a means of 
eliminating or retarding the activity of spoilage 
agents. Although low temperatures are not as 
effective in bringing about the destruction of 
spoilage agents as are high temperatures, the 
storage of perishables at low temperatures 
greatly reduces the activity of both enzymes and 
microorganisms and thereby provides a prac- 
tical means of preserving perishables in their 
original fresh state for varying periods of time. 
The degree of low temperature required for 
adequate preservation varies with the type of 
product stored and with the length of time the 
product is to be kept in storage. 

For purposes of preservation, food products 
can be grouped into two general categories : (1) 
those which are alive at the time of distribution 
and storage and (2) those which are not. Non- 
living food substances, such as meat, poultry, 
and fish, are much more susceptible to micro- 
bial contamination and spoilage than are living 
food substances, and they usually require more 
stringent preservation methods. 

With nonliving food substances, the problem 
of preservation is one of protecting dead tissue 
from all the forces of putrification and decay, 
both enzymic and microbial. In the case of 
living food substances, such as fruit and vege- 
tables, the fact of life itself affords considerable 
protection against microbial invasion, and the 
preservation problem is chiefly one of keeping 
the food substance alive while at the same time 
retarding natural enzymic activity in order to 
slow the rate of maturation or ripening. 

Vegetables and fruit are as much alive after 
harvesting as they are during the growing period. 
Previous to harvesting they receive a continuous 
supply of food substances from the growing 



130 PRINCIPLES OF REFRIGERATION 



plant, some of which is stored in the vegetable 
or fruit. After harvesting, when the vegetable or 
fruit is cut off from its normal supply of food, 
the living processes continue through utilization 
of the previously stored food substances. This 
causes the vegetable or fruit to undergo 
changes which will eventually result in deteriora- 
tion and complete decay of the product. The 
primary purpose of placing such products under 
refrigeration is to slow the living processes by 
retarding enzymic activity, thereby keeping the 
product in a preserved condition for a longer 
period. 

Animal products (nonliving food substances) 
are also affected by the activity of natural 
enzymes. The enzymes causing the most 
trouble are those which catalyze hydrolysis and 
oxidation and are associated with the break- 
down of animal fats. The principal factor 
limiting the storage life of animal products, in 
both the frozen and unfrozen states, is rancidity. 
Rancidity is caused by oxidation of animal fats 
and, since some types of animal fats are less 
stable than others, the storage life of animal 
products depends in part on fat composition. 
For example, because of the relative stability of 
beef fat, the storage life of beef is considerably 
greater than that of pork or fish whose fatty 
tissues are much less stable. 

Oxidation and hydrolysis are controlled by 
placing the product under refrigeration so that 
the activity of the natural enzymes is reduced. 
The rate of oxidation can be further reduced in 
the case of animal products by packaging the 
products in tight-fitting, gas-proof containers 
which prevent air (oxygen) from reaching the 
surface of the product. The packaging of fruit 
and vegetables in gas-proof containers, when 
stored in the unfrozen state, is not practical. 
Since these products are alive, packaging in gas- 
proof containers will cause suffocation and 
death. A dead fruit or vegetable decays very 
quickly. 

As a general rule, the lower the storage tem- 
perature, the longer is the storage life of the 
product. 

9-17. Refrigerated Storage. Refrigerated 
storage may be divided into three general cate- 
gories: (1) short-term or temporary storage, 
(2) long-term storage, and (3) frozen storage. 
For short- and long-term storage, the product is 
chilled and stored at some temperature above 



its freezing point, whereas frozen storage re- 
quires freezing of the product and storage at 
some temperature between 10° F and — 10° F, 
with 0° F being the temperature most frequently 
employed. 

Short-term or temporary storage is usually 
associated with retail establishments where 
rapid turnover of the product is normally ex- 
pected. Depending upon the product, short- 
term storage periods range from one or two days 
in some cases to a week or more in others, but 
seldom for more than fifteen days. 

Long-term storage is usually carried out by 
wholesalers and commercial storage warehouses. 
Again, the storage period depends on the type 
of product stored and upon the condition of the 
product on entering storage. Maximum storage 
periods for long-term storage range from seven 
to ten days for some sensitive products, such as 
ripe tomatoes, cantaloupes, and broccoli, and up 
to six or eight months for the more durable 
products, such as onions and some smoked 
meats. When perishable foods are to be stored 
for longer periods, they should be frozen and 
placed in frozen storage. Some fresh foods, 
however, such as tomatoes, are damaged by the 
freezing process and therefore cannot be success- 
fully frozen. When such products are to be 
preserved for long periods, some other method 
of preservation should be used. 
9-18. Storage Conditions. The optimum 
storage conditions for a product held in either 
short- or long-term storage depends upon the 
nature of the individual product, the length of 
time the product is to be held in storage, and 
whether the product is packaged or unpackaged. 
In general, the conditions required for short- 
term storage are more flexible than those 
required for long-term storage and, ordinarily, 
higher storage temperatures are permissible. 
Recommended storage conditions for both 
short- and long-term storage and the approxi- 
mate storage life for various products are listed 
in Tables 10-10 through 10-13, along with other 
product data. These data are the result of both 
experiment and experience and should be 
followed closely, particularly for long-term 
storage, if product quality is to be maintained 
at a high level during the storage period. 
9-19. Storage Temperature. Examination of 
the tables will show that the optimum storage 
temperature for most products is one slightly 



SURVEY OF REFRIGERATION APPLICATIONS 131 



above the freezing point of the product. There 
are, however, notable exceptions. 

Although the effect of incorrect storage tem- 
peratures generally is to lower product quality 
and shorten storage life, some fruits and vege- 
tables are particularly sensitive to storage tem- 
perature and are susceptible to so-called cold 
storage diseases when stored at temperatures 
above or below their critical storage tempera- 
tures. For example, citrus fruits frequently 
develop rind pitting when stored at relatively 
high temperatures. On the other hand, they are 
subject to scald (browning of the rind) and 
watery breakdown when stored at temperatures 
below their critical temperature. Bananas suffer 
peel injury when stored below 56° F, whereas 
celery undergoes soggy breakdown when stored 
at temperatures above 34° F. Although onions 
tend to sprout at temperatures above 32° F, 
Irish potatoes tend to become sweet at storage 
temperatures below 40° F. Squash, green beans, 
and peppers develop pits on their surface when 
stored at or near 32° F. Too, whereas the best 
storage temperature for most varieties of apples 
is 30° F to 32° F, some varieties are subject to 
soft scald and soggy breakdown when stored 
below 35° F. Others develop brown core at 
temperatures below 36° F, and still others 
develop internal browning when stored below 
40° F. 

9-20. Humidity and Air Motion. The 
storage of all perishables in their natural state 
(unpackaged) requires close control not only of 
the space temperature but also of space humid- 
ity and air motion. One of the chief causes of 
the deterioration of unpackaged fresh foods, 
such as meat, poultry, fish, fruit, vegetables, 
cheese, eggs, etc., is the loss of moisture from 
the surface of the product by evaporation into 
the surrounding air. This process is known as 
desiccation or dehydration. In fruit and vege- 
tables, desiccation is accompanied by shriveling 
and wilting and the product undergoes a 
considerable loss in both weight and vitamin 
content. In meats, cheese, etc., desiccation 
causes discoloration, shrinkage) and heavy 
trim losses. It also increases the rate of oxida- 
tion. Eggs lose moisture through the porous 
shell, with a resulting loss of weight and general 
downgrading of the egg. 

Desiccation will occur whenever the vapor 
pressure of the product is greater than the vapor 



pressure of the surrounding air, the rate of 
moisture loss from the product being propor- 
tional to the difference in the vapor pres- 
sures and to the amount of exposed product 
surface. 

The difference in vapor pressure between the 
product and the air is primarily a function of 
the relative humidity and the velocity of the air 
in the storage space. In general, the lower the 
relative humidity and the higher the air velocity, 
the greater will be the vapor pressure differential 
and the greater the rate of moisture loss from 
the product. Conversely, minimum moisture 
losses are experienced when the humidity in the 
storage space is maintained at a high level with 
low air velocity. Hence, 100% relative humidity 
and stagnant air are ideal conditions for pre- 
venting dehydration of the stored product. 
Unfortunately, these conditions are also con- 
ducive to rapid mold growth and the formation 
of slime (bacterial) on meats. Too, good 
circulation of the air in the refrigerated space 
and around the product is necessary for ade- 
quate refrigeration of the product. For these 
reasons, space humidity must be maintained at 
somewhat less than 100% and air velocities 
must be sufficient to provide adequate air 
circulation. The relative humidities and air 
velocities recommended for the storage of 
various products are listed in Tables 10-10 
through 10-13. 

When the product is- stored in vapor-proof 
containers, space humidity and air velocity are 
not critical. Some products, such as dried 
fruits, tend to be hydroscopic and therefore 
require storage at low relative humidities. 
9-21. Mixed Storage. Although the main- 
tenance of optimum storage conditions requires 
separate storage facilities for most products, 
this is not usually economically feasible. There- 
fore, except when large quantities of product 
are involved, practical considerations often 
demand that a number of refrigerated products 
be placed in common storage. Naturally, the 
difference in the storage conditions required by 
the various products raises a problem with 
regard to the conditions to be maintained in a 
space designed for common storage. 

As a general rule, storage conditions in such 
spaces represent a compromise and usually 
prescribe a storage temperature somewhat above 
the optimum for some of the products held in 



132 PRINCIPLES OF REFRIGERATION 



mixed storage. The higher storage tempera- 
tures are used in mixed storage in order to 
minimize the chances of damaging the more 
sensitive products which are subject to the 
aforementioned "cold storage diseases" when 
stored at temperatures below their critical 
temperature. 

Although higher storage temperatures tend 
to shorten the storage life of some of the products 
held in mixed storage, this is not ordinarily a 
serious problem when the products are stored 
only for short periods as in temporary 
storage. 

For long-term storage, most of the larger 
wholesale and commercial storage warehouses 
have a number of separate storage spaces 
available. General practice in such cases is to 
group the various products for storage, and 
only those products requiring approximately 
the same storage conditions are placed together 
in common storage. 

Another problem associated with mixed 
storage is that of odor and flavor absorption. 
Some products absorb and/or give off odors 
while in storage. Care should be taken not to 
store such products together even for short 
periods. Dairy products in particular are very 
sensitive with regard to absorbing odors and 
flavors from other products held in mixed 
storage. On the other hand, potatoes are 
probably the worst offenders in imparting off- 
flavors to other products in storage and should 
never be stored with fruit, eggs, dairy products, 
or nuts. 

9-22. Product Condition on Entering Stor- 
age. One of the principal factors determining 
the storage life of a refrigerated product is the 
condition of the product on entering storage. 
It must be recognized that refrigeration merely 
arrests or retards the natural processes of 
deterioration and therefore cannot restore to 
good condition a product which has already 
deteriorated. Neither can it make a high 
quality product out of one of initial poor 
quality. Hence, only vegetables and fruit in 
good condition should be accepted for storage. 
Those that have been bruised or otherwise 
damaged, particularly if the skin has been 
broken, have lost much of their natural protec- 
tion against microbial invasion and are there- 
fore subject to rapid spoilage by these agents. 
Too, as a general rule, since maturation and 



ripening continue after harvesting, vegetables 
and fruit intended for storage should be 
harvested before they are fully mature. The 
storage life of fully mature or damaged fruit 
and vegetables is extremely short even under 
the best storage conditions, and such products 
should be sent directly to market to avoid 
excessive losses. 

Since" a food product begins to deteriorate 
very quickly after harvesting or killing, it is 
imperative that preservation measures be taken 
immediately. To assure maximum storage life 
with minimum loss of quality, the product should 
be chilled to the storage temperature as soon 
as possible after harvesting or killing. When 
products are to be shipped over long distances 
to storage, they should be precooled and shipped 
by refrigerated transport. 
9-23. Product Chilling. Product chilling is 
distinguished from product storage in that the 
product enters the chilling room at a high 
temperature (usually either harvesting or killing 
temperature) and is chilled as rapidly as possible 
to the storage temperature, whereupon it is 
normally removed from the chilling room and 
placed in a holding cooler for storage. The 
handling of the product during the chilling 
period has a marked influence on the ultimate 
quality and storage life of the product. 

The recommended conditions for product 
chilling rooms are given in Tables 10-10 through 
10-13. Before the hot product is loaded into 
the chilling room, the chilling room temperature 
should be at the "chill finish" temperature. 
During loading and during the early part of the 
chilling period, the temperature and vapor 
pressure differential between the product and 
the chill room air will be quite large and the 
product will give off heat and moisture at a 
high rate. At this time, the temperature and 
humidity in the chill room will rise to a peak as 
indicated by the "chill start" conditions in the 
tables.* At the end of the cycle, the chill room 
temperature will again drop to the "chill finish" 
conditions. It is very important that the 
refrigerating equipment have sufficient capacity 

* The temperatures listed in the tables as chill 
start temperatures are average values and are in- 
tended for use in selecting the refrigerating equip- 
ment. Actual temperatures in the chilling room 
during the peak chilling period are usually 3° F to 
4° F higher than those listed. 



SURVEY OF REFRIGERATION APPLICATIONS 133 



to prevent the chill room temperature from 
rising excessively during the peak chilling period. 
9-24. Relative Humidity and Air Velocity 
in Chill Rooms. The importance of relative 
humidity in chilling rooms depends upon the 
product being chilled, particularly upon whether 
the product is packaged or not. Naturally, 
when the product is chilled in vapor-proof 
containers, the humidity in the chilling room 
is relatively unimportant. However, during 
loading and during the initial stages of chilling, 
chilling room humidity will be high if the con- 
tainers are wet, but will drop rapidly once the 
free moisture has been evaporated. 

Products chilled in their natural state (un- 
packaged) lose moisture very rapidly, often 
producing fog in the chilling room during the 
early stages of chilling when the product tem- 
perature and vapor pressure are high. Rapid 
chilling and high air velocity are desirable 
during this time so that the temperature and 
vapor pressure of the product are lowered as 
quickly as possible in order to avoid excessive 
moisture loss and shrinkage. High air velocity 
is needed also in order to carry away the vapor 
and thereby prevent condensation of moisture 
on the surface of the product. 

Although high air velocity tends to increase 
the rate of evaporation of moisture from the 
product, it also greatly accelerates the chilling 
rate and results in a more rapid reduction in 
product temperature and vapor pressure. Since 
the reduction in vapor pressure resulting from 
the higher chilling rate more than offsets the 
increase in the rate of evaporation occasioned 
by the higher air velocity, the net effect of the 
higher air velocity during the early stages of 
chilling is to reduce the over-all loss of moisture 
from the product. However, during the final 
stages of chilling, when the temperature and 
vapor pressure of the product are considerably 
lower, the effect of high air velocity in the chill- 
ing room is to increase the rate of moisture loss 
from the product. Therefore, the air velocity 
in the chilling room should be reduced during 
the final stages of chilling. 

As a general rule, the humidity should be 
kept at a high level when products subject to 
dehydration are being chilled. Some extremely 
sensitive products, such as poultry and fish, 
are frequently chilled in ice slush to reduce 
moisture losses during chilling. For the same 



reason, eggs are sometimes dipped in a light 
mineral oil before chilling and storage. Too, 
poultry, fish, and some vegetables are often 
packed in ice for chilling and storage. When 
products packed in ice are placed in refrigerated 
storage, the slowly melting ice keeps the surface 
of the product moist and prevents excessive 
dehydration. 

9-25. Freezing and Frozen Storage. When a 
product is to be preserved in its original fresh 
state for relatively long periods, it is usually 
frozen and stored at approximately 0° F or 
below. The list of food products commonly 
frozen includes not only those which are 
preserved in their fresh state, such as vegetables, 
fruit, fruit juices, berries, meat, poultry, sea 
foods, and eggs (not in shell), but also many 
prepared foods, such as breads, pastries, ice 
cream, and a wide variety of specially prepared 
and precooked food products, including full 
dinners. 

The factors governing the ultimate quality 
and storage life of any frozen product are: 

1 . The nature and composition of the product 
to be frozen 

2. The care used in selecting, handling, and 
preparing the product for freezing 

3. The freezing method 

4. The storage conditions. 

Only high quality products in good condition 
should be frozen. With vegetables and fruit, 
selecting the proper variety for freezing is very 
important. Some varieties are not suitable for 
freezing and will result in a low quality product 
or in one with limited keeping qualities. 

Vegetables and fruit to be frozen should be 
harvested at the peak of maturity and should be 
processed and frozen as quickly as possible 
after harvesting to avoid undesirable chemical 
changes through enzymic and microbial action. 

Both vegetables and fruit require considerable 
processing before freezing. After cleaning and 
washing to remove foreign materials — leaves, 
dirt, insects, juices, etc. — from their surfaces, 
vegetables are "blanched" in hot water or 
steam at 212° F in order to destroy the natural 
enzymes. It will be remembered that enzymes 
are not destroyed by low temperature and, 
although greatly reduced, their activity con- 
tinues at a slow rate even in food stored at 0° F 
and below. Hence, blanching, which destroys 



134 PRINCIPLES OF REFRIGERATION 




Fig. 9-2. Walk-in installation. Suspended blast 
freezer provides high-velocity air for fast freezing, 
saving valuable floor space in small areas. (Courtesy 
Carrier Corporation.) 



Fig. 9-3. Suspended blast free- 
zer applied to reach-in cabinet 
distributes blast air through 
shelves. (Courtesy Carrier 
Corporation.) 








Fig. 9-4. Freezing in one room 
and storage in another is 
accomplished by single, floor- 
mounted blast freezers. (Cour- 
tesy Carrier Corporation.) 



SURVEY OF REFRIGERATION APPLICATIONS 135 



most of the enzymes, greatly increases the 
storage life of frozen vegetables. The time 
required for blanching varies with the type and 
variety of the vegetable and ranges from 1 to l£ 
min for green beans to 1 1 min for large ears of 
corn. Although much of the microbial popula- 
tion is destroyed along with the enzymes during 
the blanching process, many bacteria survive. 
To prevent spoilage by these viable bacteria, 
vegetables should be chilled to 50° F immedi- 
ately after blanching and before they are pack- 
aged for the freezer. 

Like vegetables, fruit must also be cleaned 
and washed to remove foreign materials and 
to reduce microbial contamination. Although 
fruit is perhaps even more subject to enzymic 
deterioration than are vegetables, it is never 
blanched to destroy the natural enzymes since 
to do so would destroy the natural fresh quality 
which is so desirable. 

The enzymes causing the most concern with 
regard to frozen fruit are the ones which 
catalyze oxidation and result in rapid browning 
of the flesh. To control oxidation, fruit to be 
frozen is covered with a light sugar syrup. In 
some cases, ascorbic acid, citric acid, or sulfur 
dioxide are also used for this purpose. 

As a general rule, meat products do not 
require any special processing prior to freezing. 
However, because of consumer demand, speci- 
ally prepared meats and meat products are being 
frozen in increasing amounts. This is true also 
of poultry and sea foods. 

Because of the relative instability of their 
fatty tissue, pork and fish are usually frozen as 
soon after chilling as possible. On the other 
hand, beef is frequently "aged" in a chilling 
cooler for several days before freezing. During 
this time the beef is tenderized to some extent 
by enzymic activity. However, the aging of 
beef decreases its storage life, particularly if the 
aging period exceeds 6 or 7 days. 

With poultry, experiments indicate that 
poultry frozen within 12 to 24 hr after killing 
is more tender than that frozen immediately 
after killing. However, delaying freezing beyond 
24 hr tends to reduce storage life without 
appreciable increasing tenderness. 
9-26. Freezing Methods. Food products may 
be either sharp (slow) frozen or quick frozen. 
Sharp freezing is accomplished by placing the 
product in a low temperature room and allowing 



it to freeze slowly, usually in still air. The tem- 
perature maintained in sharp freezers ranges 
from 0° F to —40° F. Since air circulation is 
usually by natural convection, heat transfer 
from the product ranges from 3 hr to 3 days, 
depending upon the bulk of the product and 
upon the conditions in the sharp freezer. Typi- 
cal items which are sharp frozen are beef and 
pork half-carcasses, boxed poultry, panned and 
whole fish, fruit in barrels and other large 
containers, and eggs (whites, yolks, or whole) 
in 10 and 30 lb cans. 

Quick freezing is accomplished in any one or 
in any combination of three ways: (1) immer- 
sion, (2) indirect contact, and (3) air blast. 
9-27. Air Blast Freezing. Air blast freezing 
utilizes the combined effects of low temperature 
and high air velocity to produce a high rate of 
heat transfer from the product. Although the 
method employed varies considerably with the 
application, blast freezing is accomplished by 
circulating high-velocity, low-temperature air 
around the product. Regardless of the method 
used, it is important that the arrangement of 
the freezer is such that air can circulate freely 
around all parts of the product. 

Packaged blast freezers are available in both 
suspended and floor-mounted models. Typical 
applications are shown in Figs. 9-2 through 9-4. 
Blast freezing is frequently carried out in 
insulated tunnels, particularly where large 
quantities of product are to be frozen (Figs. 
9-5 and 9-6). In some instances, the product 
is carried through the freezing tunnel and 
frozen on slow-moving, mesh conveyor belts. 
The unfrozen product is placed on the con- 
veyor at one end of the tunnel and is frozen 
by the time it reaches the other end. Another 
method is to load the product on tiered dollies. 
The dollies are pushed into the tunnel and the 
product is frozen; whereupon they are pushed 
out of the freezing tunnel into a storage room 
(Fig. 9-5). 

Although blast freezing is used to freeze 
nearly all types of products, it is particularly 
suitable for freezing products of nonuniform 
or irregular sizes and shapes, such as dressed 
poultry. 

9-28. Indirect Contact Freezing. Indirect 
freezing is usually accomplished in plate freezers 
and involves placing the product on metal 
plates through which a refrigerant is circulated 



136 PRINCIPLES OF REFRIGERATION 




Fig. 9-5. Packaged blast freezers 
applied to freezing tunnel. High 
velocity, — 15° F air is blasted 
through trucks. (Courtesy Car- 
rier Corporation.) 



(Fig. 9-7). Since the product is in direct thermal 
contact with the refrigerated plate, heat transfer 
from the product occurs primarily by conduction 
so that the efficiency of the freezer depends, for 
the most part, on the amount of contact surface. 
This type of freezer is particularly useful when 
products are frozen in small quantities. 

One type of plate freezer widely used by the 
larger commercial freezers to handle small, 
flat, rectangular, consumer-size packages is the 
multiplate freezer. The multiplate freezer 
consists of a series of horizontal, parallel, 
refrigerated plates which are actuated by 
hydraulic pressure so that they can be opened 
to receive the product between them and then 
closed on the product with any desired pressure. 
When the plates are closed, the packages are 
held tightly between the plates. Since both the 
top and the bottom of the packages are in good 
thermal contact with the refrigerated plates, 
the rate of heat transfer is high and the product 
is quickly frozen. 

9-29. Immersion Freezing. Immersion freez- 
ing is accomplished by immersing the product in 
a low temperature brine solution, usually either 
sodium chloride or sugar. Since the refrigerated 
liquid is a good conductor and is in good 
thermal contact with all the product, heat 
transfer is rapid and the product is completely 
frozen in a very short time. 

Another advantage of immersion freezing is 
that the product is frozen in individual units 
rather than fused together in a mass. 

The principal disadvantage of immersion 



freezing is that juices tend to be extracted from 
the product by osmosis. This results in con- 
tamination and weakening of the freezing 
solution. Too, where a sodium chloride brine 
is used, salt penetration into the product may 
sometimes be excessive. On the other hand, 
when fruit is frozen in a sugar solution, sugar 
penetration into the fruit is entirely beneficial. 

The products most frequently frozen by 
immersion are fish and shrimp. Immersion is 
particularly suitable for freezing fish and 
shrimp at sea, since the immersion freezer is 
relatively compact and space aboard ship is at 
a premium. In addition, immersion freezing 
produces a "glaze" (thin coating of ice) on the 
surface of the product which helps to prevent 
dehydration of unpackaged products during 
the storage period. 

9-30. Quick Freezing vs. Sharp Freezing. 
Quick frozen products are nearly always 
superior to those which are sharp (slow) frozen. 
D. K. Tressler, in 1932, summarized the views 
of R. Plank, H. F. Taylor, C. Birdseye, and 
G. A. Fitzgerald, and stated the following as the 
main advantages of quick freezing over slow 
freezing: 

1 . The ice crystals formed are much smaller, 
and therefore cause much less damage to cells. 

2. The freezing period being much shorter, 
less time is allowed for the diffusion of salts and 
the separation of water in the form of ice. 

3. The product is quickly cooled below the 
temperature at which bacterial, mold, and yeast 



SURVEY OF REFRIGERATION APPLICATIONS 137 




I 

e 



o 
u 



i 



V 

L. 

& 



I 



138 PRINCIPLES OF REFRIGERATION 




Fig. 9-7. Plate freezer for indirect 
contact freezing. (Courtesy Dole 
Refrigerating Company.) 



growth occurs, thus preventing decomposition 
during freezing.* 

The principal difference between quick freez- 
ing and sharp freezing is in the size, number, 
and location of the ice crystals formed in the 
product as cellular fluids are solidified. When 
a product is slow frozen, large ice crystals are 
formed which result in serious damage to the 
tissue of some products through cellular break- 
down. Quick freezing, on the other hand, 
produces smaller ice crystals which are formed 
almost entirely within the cell so that cellular 
breakdown is greatly reduced. Upon thawing, 
products which have experienced considerable 
cellular damage are prone to lose excessive 
amounts of fluids through "drip" or "bleed," 
with a resulting loss of quality. 

Ice-crystal formation begins in most products 
at a temperature of approximately 30° F and, 
although some extremely concentrated fluids 
still remain unfrozen even at temperatures below 
—50° F, most of the fluids are solidified by the 

* Air Conditioning Refrigerating Data Book, 
Applications Volume, 5th Edition, American Society 
of Refrigerating Engineers, 1954-55, p. 1-02. 



time the product temperature is lowered to 
25° F. The temperature range between 30° F 
and 25° F is often referred to as the zone of 
maximum ice-crystal formation, and rapid heat 
removal through this zone is desirable from the 
standpoint of product quality. This is particu- 
larly true for fruits and vegetables because both 
undergo serious tissue damage when slow frozen. 
Since animal tissue is much tougher and 
much more elastic than plant tissue, the freezing 
rate is not as critical in the freezing of meats 
and meat products as it is in fruits and vege- 
tables. Recent experiment indicates that poultry 
and fish suffer little, if any, cellular damage 
when slow frozen. This does not mean, how- 
ever, that quick frozen meats are not superior 
to those which are slow frozen, but only that, 
for the standpoint of cellular damage, quick 
freezing is not as important in the freezing of 
meats as it' is in fruits and vegetables. For 
example, poultry that is slow frozen takes on a 
darkened appearance which makes it much less 
attractive to the consumer. This alone is 
enough to justify the quick freezing of poultry. 
Too, in all cases, quick freezing reduces the 
processing time and, consequently, the amount 



SURVEY OF REFRIGERATION APPLICATIONS 139 



of bacterial deterioration. This is especially 
worthwhile in the processing of fish because of 
their tendency to rapid spoilage. 
9-31. Packaging Materials. Dehydration, one 
of the principal factors limiting the storage life 
of frozen foods, is greatly reduced by proper 
packaging. Unpackaged products are subject 
to serious moisture losses not only during the 
freezing process but also during the storage 
period. While in storage, unpackaged frozen 
products lose moisture to the air continuously 
by sublimation. This eventually results in a 
condition known as "freezer-burn," giving the 
product a white, leathery appearance. Freezer- 
burn is usually accompanied by oxidation, 
flavor changes, and loss of vitamin content. 

With few exceptions, all products are pack- 
aged before being placed in frozen storage. 
Although most products are packaged before 
freezing, some, such as loose frozen peas and lima 
beans, are packaged after the freezing process. 

To provide adequate protection against 
dehydration and oxidation, the packaging 
material should be practically 100% gas and 
vapor proof and should fit tightly around the 
product to exclude as much air as possible. Too, 
air spaces in packages have an insulating effect 
which reduce the freezing rate and increase 
freezing costs. 

The fact that frozen products are in compe- 
tition to products preserved by other methods 
introduces several factors which must be taken 
into account when selecting packaging materials. 
When the product is to be sold directly to the 
consumer, the package must be attractive and 
convenient to use in order to stimulate sales. 
From a cost standpoint, the package should be 
relatively inexpensive and of such a nature that 
it permits efficient handling so as to reduce 
processing costs. 

Some packaging materials in general use are 
aluminum foil, tin cans, impregnated paper- 
board cartons, paper-board cartons over- 
wrapped with vapor-proof wrappers, wax paper, 
cellophane, polyethylene, and other sheet 
plastics. 

Frozen fish are often given an ice glaze (a 
thin coating of ice) which provides an excellent 
protective covering. However, since the ice 
glaze is very brittle, glazed fish must be handled 
carefully to avoid breaking the glaze. Too, since 
the ice glaze gradually sublimes to the air, the 



fish must be reglazed approximately once a 
month by dipping into fresh water or by 
spraying. 

9-32. Frozen Storage. The exact temperature 
required for frozen storage is not critical, 
provided that it is sufficiently low and that it 
does not fluxuate. Although 0°F is usually 
adequate for short-term (retail) storage, -5° F 
is the best temperature for all-around long-term 
(wholesale) storage. When products having 
unstable fats (oxidizable, free, fatty acids) are 
stored in any quantity, the storage temperature 
should be held at —10° F or below in order to 
realize the maximum storage life. 

When products are stored above — 20° F, 
which is normally the case, the temperature of 
the storage room should be maintained constant 
with a variation of not more than 1 ° F in either 
direction. Variations in storage temperature 
cause alternate thawing and refreezing of some 
of the juices in the product. This tends to 
increase the size of the ice crystals in the 
product and eventually results in the same 
type of cellular damage as occurs with slow 
freezing. 

Since many packaging materials do not offer 
complete protection against dehydration, the 
relative humidity should be kept at a high level 
(85 % to 90%) in frozen storage rooms, particu- 
larly for long-term storage. 

Proper stacking of the product is also 
essential. Stacking should always be such that 
it permits adequate air circulation around the 
product. It is particularly important to leave a 
good size air space between the stored product 
and the walls of the storage room. In addition 
to permitting air circulation around the product, 
this eliminates the possibility of the product 
absorbing heat directly from the warm walls. 
9-33. Commercial Refrigerators. The term 
"commercial refrigerator" is usually applied to 
the smaller, ready-built, refrigerated fixtures of 
the type used by retail stores and markets, 
hotels, restaurants, and institutions for the 
processing, storing, displaying, and dispensing 
of perishable commodities. The term is some- 
times applied also to the larger, custom-built 
refrigerated fixtures and rooms used for these 
purposes. 

Although there are a number of special 
purpose refrigerated fixtures which defy classi- 
fication, in general, commercial fixtures can 



140 PRINCIPLES OF REFRIGERATION 




Fig. 9-8. Typical reach-in refrigera- 
tor. (Courtesy Tyler Refrigeration 
Corporation.) 



be grouped into three principal categories: 
(1) reach-in refrigerators, (2) walk-in coolers, 
and (3) display cases. 

9-34. Reach-In Refrigerators. The reach-in 
refrigerator is probably the most versatile and 
the most widely used of all commercial fixtures. 
Typical users are grocery stores, meat markets, 
bakeries, drug stores, lunch counters, res- 
taurants, florists, hotels, and institutions of all 
kinds. Whereas some reach-in refrigerators 
serve only a storage function, others are used 
for both storage and display (Fig. 9-8). Those 
serving only the storage function usually have 
solid doors, whereas those used for display 
have glazed doors. 

9-35. Walk-In Coolers. Walk-in coolers are 
primarily storage fixtures and are available in a 
wide variety of sizes to fit every need. Nearly 
all retail stores, markets, hotels, restaurants, 
institutions, etc., of any size employ one or more 
walk-in coolers for the storage of perishables 
of all types. Some walk-in coolers are equipped 
with glazed reach-in doors. This feature is 
especially convenient for the storing, displaying, 
and dispensing of such items as dairy products, 
eggs, and beverages. Walk-in coolers with 
reach-in doors are widely used in grocery stores, 
particularly drive-in groceries, for handling 
such items. 

9-36. Display Cases. The principal function of 
any kind of display fixture is to display the 



product or commodity as attractively as possible 
in order to stimulate sales. Therefore, in the 
design of refrigerated display fixtures, first 
consideration is given to the displaying of the 
product. In many cases, this is not necessarily 
compatible with providing the optimum storage 
conditions for the product being displayed. 
Hence, the storage life of a product in a display 
fixture is frequently very limited, ranging from 
a few hours in some instances to a week or more 
in others, depending upon the type of product 
and upon the type of fixture. 

Display fixtures are of two general types: 
(1) the self-service case, from which the customer 
serves himself directly, and (2) the service case, 
from which the customer is usually served by 
an attendant. The former is very popular in 
supermarkets and other large, retail, self-service 
establishments, whereas the service case finds 
use in the smaller groceries, markets, bakeries, 
etc. Typical service cases are shown in Figs. 9-9 
and 9-10. 

Self-service cases are of two types, open and 
closed, with the open type gaining rapidly in 
popularity. With the advent of the super- 
market, the trend has been increasingly toward 
the open type self-service case, and the older, 
closed type self-service cases are becoming 
obsolete. Several of the more popular types of 
open self-service cases are shown in Figs. 9-11 
and 9-12. These are used to display meat 



SURVEY OF REFRIGERATION APPLICATIONS 141 



vegetables, fruit, frozen foods, ice cream, dairy 
products, delicatessen items, etc. The design 
of the case varies somewhat with the particular 
type of product being displayed. Too, designs 
are available for both wall and island installa- 
tion. Although some provide additional storage 
space, others do not. 

9-37. Special Purpose Fixtures. Although all 
the refrigerated fixtures discussed in the preced- 
ing sections are available in a variety of designs 
in order to satisfy the specific requirements of 
individual products and applications, a number 
of special purpose fixtures is manufactured 
which may or may not fall into one of the three 
general categories already mentioned. Some of 
the more common special purpose fixtures are 




Fig. 9-9. Conventional single-duty service case for 
displaying meats. (Courtesy Tyler Refrigeration 
Corporation.) 

beverage coolers, milk coolers (dairy farm), milk 
and beverage dispensers, soda fountains, ice 
cream makers, water coolers, ice makers, back- 
bar refrigerators, florist boxes, dough retarders, 
candy cases, and mortuary refrigerators. 
9-38. Frozen Food Locker Plants. Normally, 
the function of a frozen food locker plant is to 
process and freeze foods for individual families 
and other groups, either for take-home storage 
or for storage at the locker plant. When 
storage is at the plant, the customer rents a 
storage space (locker) and calls at the plant for 
one or more packages as needed. 




Fig. 9-10. Double-duty service case for displaying 
meats. (Courtesy Tyler Refrigeration Corporation.) 



As a general rule, a locker plant furnishes all 
or most of the following facilities and/or 
services: 

1. A chilling room for chilling freshly killed 
meats. 

2. A cold storage room for holding products 
under refrigeration while awaiting preparation 
and processing prior to freezing. 

3. A processing room where the products are 
processed and packaged for the freezer. 




Fig. 9-11. High multishelf produce sales 
(Courtesy Tyler Refrigeration Corporation.) 



142 PRINCIPLES OF REFRIGERATION 




Fig. 9-12. Open-type display case for 
frozen foods and icecream. (Courtesy 
Tyler Refrigeration Corporation.) 



4. A freezing room or cabinet in which the 
food is frozen prior to being placed in storage. 

5. A low temperature room containing the 
storage lockers. 

6. A low temperature bulk storage room. 

7. An aging room where certain meats are 
kept under refrigeration and allowed to age 
(tenderize) for periods usually ranging from 
7 to 10 days. 



8. A curing and smoking room for handling 
bacon, ham, sausage, and other cured meats. 

Services such as slaughtering, lard rendering, 
sausage making, etc., are also provided by some 
plants. 

The layout of a typical locker plant is shown 
in Fig. 9-13. The recommended design condi- 
tions for the various spaces in the locker plant 
are given in Fig. 9-14. The average size of the 



v//>y////Y//////^/Y//w//y/^/// , /yy/s/y/> , M^ 




Fig. 9-13. Typical locker plant. (ASRE Data Book, Applications Volume, 1956-57.) Reproduced by permission 
of American Society of Heating, Refrigerating, and Air-Conditioning Engineers. 



Type of space 



SURVEY OF REFRIGERATION APPLICATIONS 143 
Locker Plant Design Conditions 

Refrigerant Insulation 



Room 
temperature 



temperature thickness, inches. 



Work room, process room, 








and kitchen 




Atmospheric 


None 


None 


Chill room 




34 to 36 F Design for 

35 F 


20 to 25 F below room 
temperature, for gravity 
circulation; 10 to 15 F 
below room tempera- 
ture for forced air cir- 
culation 


3 to 8 


Aging room 




34 to 36 F Design for 

35 F 


Same as chill room 


3 to 8 


Curing room 




38 to 40 F Design for 
40F 


Same as chill room 


3 to 8 


Freezing room 


(gravity 


-10 to -20 F 


-20 to -30 F 


6 to 12 


air circulation) 










Freezer cabinet (in locker 


Not important 


-15 to -20 F 


1 or 2 


room) 










Blast freezer 




Depends on type of 
system used 


-10 to -15 F 


6 to 12 


Locker room or 


bulk 


OF 


-15 to -20 F 


6 to 12 


storage 











Fig. 9-14. (ASRE Data Book, Applications Volume, 1956-57. 
of Heating, Refrigerating and Air Conditioning Engineers.) 



Reproduced by permission of American Society 



individual locker is 6 cu ft and the average 
product storage capacity is approximately 35 
to 40 lb per cubic foot. Minimum product 
turnover is approximately 2 lb per locker per 
day. Standard practice is to base chilling room 
and freezer capacities on the handling of 2 to 4 
lb of product per locker per day. 
9-39. Summary. Recognizing that a thorough 
knowledge of the application itself is a pre- 
requisite to good system design and proper 
equipment selection, we have devoted the mate- 
rial in this chapter to a brief survey of a few of 
the applications of mechanical refrigeration, 
with special emphasis being given to the area 
of commercial refrigeration. 

Obviously, the applications of mechanical 
refrigeration are too many and too varied to 
permit detailed consideration of each and 
every type. Fortunately, this is neither necessary 
nor desirable since methods of system designing 
and equipment selection are practically the 



same for all types of applications. Commerical 
refrigeration was selected for emphasis because 
this area embraces a wide range of applications 
and because the problems encountered in this 
area are representative of those in the other 
areas. Hence, even though the discussion in 
this chapter and in those which follow deals 
chiefly with commercial refrigeration, the 
principals of system design and the methods of 
equipment selection developed therein may be 
applied to all types of mechanical refrigeration 
applications. 

Although no attempt is made in this book 
to discuss air conditioning as such except in a 
very general way, it should be pointed out that 
most commercial refrigeration applications, 
particularly those concerned with product 
storage, involve air conditioning in that they 
ordinarily include close control of the tempera- 
ture, humidity, motion, and cleanliness of the air 
in the refrigerated space. 



10 

Cooling Load 
Calculations 



10-1. The Cooling Load. The cooling load on 
refrigerating equipment seldom results from 
any one single source of heat. Rather, it is the 
summation of the heat which usually evolves 
from several different sources. Some of the 
more common sources of heat which supply 
the load on refrigerating equipment are: 

1. Heat that leaks into the refrigerated space 
from the outside by conduction through the 
insulated walls. 

2. Heat that enters the space by direct 
radiation through glass or other transparent 
materials. 

3. Heat that is brought into the space by 
warm outside air entering the space through 
open doors or through cracks around windows 
and doors. 

4. Heat given off by a warm product as its 
temperature is lowered to the desired level. 

5. Heat given off by people occupying the 
refrigerated space. 

6. Heat given off by any heat-producing 
equipment located inside the space, such as 
electric motors, lights, electronic equipment, 
steam tables, coffee urns, hair driers, etc. 

The importance of any one of these heat 
sources with relation to the total cooling load 
on the equipment varies with the individual 
application. Not all them will be factors in 
every application, nor will the cooling load in 
any one application ordinarily include heat 
from all these sources. However, in any given 



application, it is essential that consideration be 
given to all heat sources present and that all the 
heat evolving from them be taken into account 
in the over-all calculation. 
10-2. Equipment Running Time. Although 
refrigerating equipment capacities are normally 
given in Btu per hour, in refrigeration applica- 
tions the total cooling load is usually calculated 
for a 24-hr period, that is, in Btu per 24 hr. 
Then, to determine the required Btu per hour 
capacity of the equipment, the total load for the 
24-hr period is divided by the desired running 
time for the equipment, viz: 

RequiredBtu/hr Totalcoolingload)Btu/24hr 

equipment = 2 : '. 

capacity Desired running time 

(10-1) 

Because of the necessity for defrosting the 
evaporator at frequent intervals, it is not 
practical to design the refrigerating system in 
such a way that the equipment must operate 
continuously in order to handle the load. In 
most cases, the air passing over the cooling coil 
is chilled to a temperature below its dew point 
and moisture is condensed out of the air onto 
the surface of the cooling coil. When the tem- 
perature of the coil surface is above the freezing 
temperature of water, the moisture condensed 
out of the air drains off the coil into the con- 
densate pan and leaves the space through the con- 
densate drain. However, when the temperature of 
the cooling coil is below the freezing tempera- 
ture of water, the moisture condensed out of 
the air freezes into ice and adheres to the surface 
of the coil, thereby causing "frost" to accumu- 
late on the coil surface. Since frost accumulation 
on the coil surface tends to insulate the coil 
and reduce the coil's capacity, the frost must 
be melted off periodically by raising the surface 
temperature of the coil above the freezing point 
of water and maintaining it at this level until 
the frost has melted off the coil and left the 
space through the condensate drain. 

No matter how the defrosting is accomplished, 
the defrosting requires a certain amount of 
time, during which the refrigerating effect of the 
system must be stopped. 

One method of defrosting the coil is to stop 
the compressor and allow the evaporator to 
warm up to the space temperature and remain 
at this temperature for a sufficient length of 



144 



COOLING LOAD CALCULATIONS 145 



time to allow the frost accumulation to melt off 
the coil. This method of defrosting is called 
"off-cycle" defrosting. Since the heat required 
to melt the frost in off-cycle defrosting must 
come from the air in the refrigerated space, 
defrosting occurs rather slowly and a consider- 
able length of time is required to complete the 
process. Experience has shown that when off- 
cycle defrosting is used, the maximum allowable 
running time for the equipment is 16 hr out of 
each 24-hr period, the other 8 hr being allowed 
for the defrosting. This means, of course, that 
the refrigerating equipment must have sufficient 
capacity to accomplish the equivalent of 24 hr of 
cooling in 1 6 hr of actual running time. Hence, 
when off-cycle defrosting is used, the equipment 
running time used in Equation 10-1 is approxi- 
mately 16 hr. 

When the refrigerated space is to be main- 
tained at a temperature below 34° F, off-cycle 
defrosting is not practical. The variation in 
space temperature which would be required in 
order to allow the cooling coil to attain a 
temperature sufficiently high to melt off the 
frost during every off cycle would be detrimental 
to the stored product. Therefore where the 
space temperature is maintained below 34° F, 
some method of automatic defrosting is 
ordinarily used. In such cases the surface of the 
coil is heated artificially, either with electric 
heating elements, with water, or with hot gas 
from the discharge of the compressor (see 
Chapter 20). 

Defrosting by any of these means is accom- 
plished much more quickly than when off-cycle 
defrosting is used. Hence, the off-cycle time 
required is less for automatic defrosting and 
the maximum allowable running time for the 
equipment is greater than for the aforementioned 
off-cycle defrosting. For systems using auto- 
matic defrosting the maximum allowable 
running time is from 18 to 20 hr out of each 
24-hr period, depending upon how often de- 
frosting is necessary for the application in 
question. As a general rule, the 18 hr running 
time is used. 

It is of interest to note that since the tempera- 
ture of the cooling coil in comfort air condition- 
ing applications is normally around 40° F, no 
frost accumulates on the coil surface and, there- 
fore, no off-cycle time is required for defrosting. 
For this reason, air conditioning systems are 



usually designed for continuous run and cooling 
loads for air conditioning applications are 
determined directly in Btu per hour. 
10-3. Cooling Load Calculations. To simplify 
cooling load calculations, the total cooling load 
is divided into a number of individual loads 
according to the sources of heat supplying the 
load. The summation of these individual loads 
is the total cooling load on the equipment. 

In commercial refrigeration, the total cooling 
load is divided into four separate loads, viz: 
(1) the wall gain load, (2) the air change load, 
(3) the product load, and (4) the miscellaneous 
or supplementary load. 

10-4. The Wall Gain Load, The wall gain 
load, sometimes called the wall leakage load, 
is a measure of the heat which leaks through the 
walls of the refrigerated space from the outside 
to the inside. Since there is no perfect insulation, 
there is always a certain amount of heat passing 
from the outside to the inside whenever the 
inside temperature is below that of the outside. 
The wall gain load is common to all refrigeration 
applications and is ordinarily a considerable 
part of the total cooling load. Some exceptions 
to this are liquid chilling applications, where the 
outside area of the chiller is small and the walls 
of the chiller are well insulated. In such cases, 
the leakage of heat through the walls of the 
chiller is so small in relation to the total cooling 
load that its effect is negligible and it is usually 
neglected. On the other hand, commercial 
storage coolers and residential air conditioning 
applications are both examples of applications 
wherein the wall gain load usually accounts for 
the greater portion of the total load. 
10-5. The Air Change Load. When the door 
of a refrigerated space is opened, warm outside 
air enters the space to replace the more dense 
cold air which is lost from the refrigerated space 
through the open door. The heat which must 
be removed from this warm outside air to 
reduce its temperature to the space temperature 
becomes a part of the total cooling load on the 
equipment. This part of the total load is called 
the air change load. 

The relationship of the air change load to the 
total cooling load varies with the application. 
Whereas in some applications the air change 
load is not a factor at all, in others it represents 
a considerable portion of the total load. For 
example, with liquid chillers, there are no doors 



146 PRINCIPLES OF REFRIGERATION 



or other openings through which air can pass 
and therefore the air change load is nonexistent. 
On the other hand, the reverse is true for air 
conditioning applications, where, in addition 
to the air changes brought about by door 
openings, there is also considerable leakage of 
air into the conditioned space through cracks 
around windows and doors and in other parts 
of the structure. Too, in many air conditioning 
applications outside air is purposely introduced 
into the conditioned space to meet ventilating 
requirements. When large numbers of people 
are in the conditioned space, the quantity of 
fresh air which must be brought in from the 
outside is quite large and the cooling load 
resulting for the cooling of this air to the tem- 
perature of the conditioned space is often a 
large part of the total cooling load in such 
applications. 

In air conditioning applications, the air change 
load is called either the ventilating load or the 
infiltration load. The term ventilating load is 
used when the air changes in the conditioned 
space are the result of deliberate introduction 
of outside air into the space for ventilating 
purposes. The term infiltration load is used 
when the air changes are the result of the 
natural infiltration of air into the space through 
cracks around windows and doors. Every air 
conditioning application will involve either an 
infiltration load or a ventilating load, but never 
both in the same application. 

Since the doors on commercial refrigerators 
are equipped with well-fitted gaskets, the cracks 
around the doors are tightly sealed and there is 
little, if any, leakage of air around the doors of a 
commercial fixture in good condition. Hence, 
in commercial refrigeration, the air changes are 
usually limited to those which are brought about 
by actual opening and closing of the door or 
doors. 

10-6. The Product Load. The product load is 
made up of the heat which must be removed 
from the refrigerated product in order to reduce 
the temperature of the product to the desired 
level. The term product as used here is taken 
to mean any material whose temperature is 
reduced by the refrigerating equipment and 
includes not only perishable commodities, such 
as foodstuff, but also such items as welding 
electrodes, masses of concrete, plastic, rubber, 
and liquids of all kinds. 



The importance of the product load in relation 
to the total cooling load, like all others, varies 
with the application. Although it is nonexistent 
in some applications, in others it represents 
practically the entire cooling load. Where the 
refrigerated cooler is designed for product 
storage, the product is usually chilled to the 
storage temperature before being placed in the 
cooler and no product load need be considered 
since the product is already at the storage 
temperature. However, in any instance where 
the product enters a storage cooler at a tem- 
perature above the storage temperature, the 
quantity of heat which must be removed from 
the product in order to reduce its temperature 
to the storage temperature must be considered 
as a part of the total load on the cooling 
equipment. 

In some few instances, the product enters the 
storage fixture at a temperature below the 
normal storage temperature for the product. 
A case in point is ice cream which is frequently 
chilled to a temperature of 0°F or -10° F 
during the hardening process, but is usually 
stored at about 10° F, which is the ideal dipping 
temperature. When such a product enters 
storage at a temperature below the space 
temperature, it will absorb heat from the 
storage space as it warms up to the storage 
temperature and thereby produce a certain 
amount of refrigerating effect of its own. In 
other words, it provides what might be termed 
a negative product load which could theoreti- 
cally be subtracted from the total cooling load. 
This is never done, however, since the refriger- 
ating effect produced is small and is not 
continuous in nature. 

The cooling load on the refrigerating equip- 
ment resulting from product cooling may be 
either intermittent or continuous, depending 
on the application. The product load is a part 
of the total cooling load only while the tempera- 
ture of the product is being reduced to the 
storage temperature. Once the product is 
cooled to the storage temperature, it is no 
longer a source of heat and the product load 
ceases to be a part of the load on the equipment. 
An exception to this is in the storage of fruit 
and vegetables which give off respiration heat 
for the entire time they are in storage even 
though there is no further decrease in their 
temperature (see Section 10-17). 



COOLING LOAD CALCULATIONS 147 



There are, of course, a number of refrigera- 
tion applications where product cooling is 
more or less continuous, in which case the 
product load is a continuous load on the equip- 
ment. This is true, for instance, in chilling 
coolers where the primary function is to chill 
the warm product to the desired storage tem- 
perature. When the product has been cooled 
to the storage temperature, it is usually moved 
out of the chilling room into a storage room and 
the chilling room is then reloaded with warm 
product. In such cases, the product load is 
continuous and is usually a large part of the 
total load on the equipment. 

Liquid chilling is another application wherein 
the product provides a continuous load on the 
refrigerating equipment. The flow of the liquid 
being chilled through the chiller is continuous 
with warm liquid entering the chiller and cold 
liquid leaving. In this instance, the product 
load is practically the only load on the equip- 
ment since there is no air change load and the 
wall gain load is negligible, as is the miscel- 
laneous load. 

In air conditioning applications there is no 
product load as such, although there is often a 
"pull-down load," which, in a sense, may be 
thought of as a product load. 
10-7. The Miscellaneous Load. The miscel- 
laneous load, sometimes referred to as the 
supplementary load, takes into account all 
miscellaneous sources of heat. Chief among 
these are people working in or otherwise 
occupying the refrigerated space along with 
lights or other electrical equipment operating 
inside the space. 

In most commercial refrigeration applica- 
tions the miscellaneous load is relatively small, 
usually consisting only of the heat given off by 
lights and fan motors used inside the space. 

In air conditioning applications, there is no 
miscellaneous load as such. This is not to say 
that human occupancy and equipment are not 
a part of the cooling load in air conditioning 
applications. On the contrary, people and 
equipment are often such large factors in the 
air conditioning load that they are considered 
as separate loads and are calculated as such. 
For example, in those air conditioning applica- 
tions where large numbers of people occupy the 
conditioned space, such as churches, theaters, 
restaurants, etc., the cooling load resulting from 



human occupancy is frequently the largest single 
factor in the total load. Too, many air condi- 
tioning systems are installed for the sole purpose 
of cooling electrical, electronic, and other types 
of heat-producing equipment. In such cases, 
the equipment usually supplies the greater 
portion of the cooling load. 
10-8. Factors Determining the Wall Gain 
Load. The quantity of heat transmitted through 
the walls of a refrigerated space per unit of time 
is the function of three factors whose relation- 
ship is expressed in the following equation: 



Q=A x U x D 



(10-2) 



where Q = the quantity of heat transferred in 
Btu/hr 
A — the outside surface area of the wall 

(square feet) 
U = the over-all coefficient of heat trans- 
mission (Btu/hr/sq ft/° F) 
D = the temperature differential across 
the wall (° F) 

The coefficient of transmission or "U" factor 
is a measure of the rate at which heat will pass 
through a 1 sq ft area of wall surface from the 
air on one side to the air on the other side for 
each 1° F of temperature difference across the 
wall. The value of the U factor is given in 
Btu per hour and depends on the thickness of 
the wall and on the materials used in the wall 
construction. Since it is desirable to prevent as 
much heat as possible from entering the space 
and becoming a load on the cooling equipment, 
the materials used in the construction of cold 
storage walls should be good thermal insulators 
so that the value of U is kept as low as is 
practical. 

According to Equation 10-2, once the U 
factor is established for a wall, the rate of heat 
flow through the wall varies directly with the 
surface area of the wall and with the temperature 
differential across the wall. Since the value of 
U is given in Btu/hr/sq ft/° F, the total quantity 
of heat passing through any given wall in 1 hr 
can be determined by multiplying the U factor 
by the wall area in square feet and by the tem- 
perature difference across the wall in degrees 
Fahrenheit, that is, by application of Equation 
10-2. 

Example 10-1. Determine the total quan- 
tity of heat in Btu per hour which will pass 
through a wall 10 ft by 20 ft, if the U factor 



148 PRINCIPLES OF REFRIGERATION 

for the wall is 0.16 Btu/hr/sq ft/ F and the 
temperature on one side of the wall is 40° F 
while the temperature on the other side is 95° F. 

Solution 
Total wall area 

Temperature differ- 
ential across wall, ° F 

Applying Equation 
10-2, the heat gain 
through the wall 



■ 10 ft x 20 ft 
= 200 sq ft 

. 95° - 40° 
>55°F 

200 x 0.16 x 55 
1760 Btu/hr 



Since the value of U in Equation 10-2 is in 
Btu per hour, the result obtained from Equation 
10-2 is in Btu per hour. To determine the wall 
gain load in Btu per 24 hr as required in refriger- 
ation load calculations, the result of Equation 
10-2 is multiplied by 24 hr. Hence, for calcula- 
tion cooling loads in refrigeration applications, 
Equation 10-2 is written to include this multi- 
plication, viz: 



Q=AxUxDx24 



(10-3) 



10-9. Determination of the (/Factor. Over- 
all coefficients of transmission or U factors have 
been determined for various types of wall 
construction and these values are available in 
tabular form. Tables 10-1 through 10-3 list U 
values for various types of cold storage walls. 

Example 10-2. From Table 10-1, determine 
the U factor for a wall constructed of 6-in. clay 
tile with 4 in. of corkboard insulation. 

Solution. Turn to Table 10-1 and select the 
appropriate type of wall construction (third 



Air spaces 




Concrete 
aggregate 



Fig. 10-1. Concrete aggregate building block. 



from top). In the next column select the desired 
thickness of clay tile (6 in.) and move to the 
right to the column listing values for 4 in. of 
insulation. Read the U factor of the wall, 0.064 
Btu/hr/sq ft/° F. 

Should it be necessary, the U factor for any 
type of wall construction can be readily calcu- 
lated provided that either the conductivity or the 
conductance of each of the materials used in the 
wall construction is known. The conductivity 
or conductance of most of the materials used in 
wall construction can be found in tables. Too, 
this information is usually available from the 
manufacturer or producer of the material. 
Table 10-4 lists the thermal conductivity or the 
conductance of materials frequently used in the 
construction of cold storage walls. 

The thermal conductivity or k factor of a 
material is the rate in Btu per hour at which 
heat passes through a 1 sq ft cross section of the 
material 1 in. thick for each 1 ° F of temperature 
difference across the material. 

Whereas the thermal conductivity or k factor 
is available only for homogeneous materials 
and the value given is always for a 1 in. thickness 
of the material, the thermal conductance or C 
factor is available for both homogeneous and 
nonhomogeneous materials and the value given 
is for the specified thickness of the material. 

For any homogeneous material, the thermal 
conductance can be determined for any given 
thickness of the material by dividing the k 
factor by the thickness in inches. Hence, for a 
homogeneous material, 

C=\ (10-4) 

x 

where x = the thickness of material in inches. 

Example 10-3. Determine the thermal 
conductance for a 5 in. thickness of corkboard. 

Solution 
From Table 10-4, 
k factor of cork- 
board = 0.30 Btu/hr/sq ft/in/° F 

Applying Q , 

Equation 10-4, C =_ 

= 0.06 Btu/hr/sq ft/° F 

Since the rate of heat transmission through 
nonhomogeneous materials, such as the concrete 
building block in Fig. 10-1, will vary in the 



COOLING LOAD CALCULATIONS 149 



several parts of the material, the C factor from 
nonhomogeneous materials must be determined 
by experiment. 

The resistance that a wall or a material offers 
to the flow of heat is inversely proportional to 
the ability of the wall or material to transmit 
heat. Hence, the over-all thermal resistance of 
a wall can be expressed as the reciprocal of the 
over-all coefficient of transmission, whereas the 
thermal resistance of an individual material 
can be expressed as the reciprocal of its conduc- 
tivity or conductance, viz: 



Over-all thermal resist- 
ance 

Thermal resistance of 
an individual material 



~U 

1 1 x 

-l or c or k 



The terms I Ik and 1/C express the resistance 
to heat flow through a single material from 
surface to surface only and do not take into 
account the thermal resistance of the thin film 
of air which adheres to all exposed surfaces. In 
determining the over-all thermal resistance to 
the flow of heat through a wall from the air on 
one side to the air on the other side, the resist- 
ance of the air on both sides of the wall should 
be considered. Air film coefficients or surface 
conductances for average wind velocities are 
given in Table 10-5/*. 

When a wall is constructed of several layers 
of different materials the total thermal resistance 
of the wall is the sum of the resistances of the 
individual materials in the wall construction, 
including the air films, viz: 

\ \ x x x \ 

U fi **1 *2 k n Jo 



Therefore 
U * 



1 



I X X x 1 

ft k x k t k n f„ 



where rr = surface conductance of inside wall, 

' ' floor, or ceiling 

1 

— = surface conductance of outside wall, 

■* * floor, or roof 

Note. When nonhomogeneous materials are 
used, 1/C is substituted for xjk. 



Example 10-4. Calculate the value of £/for 
a wall constructed of 8 in. cinder aggregate 
building blocks, insulated with 4 in. of cork- 
board, and finished on the inside with 0.5 in. of 
cement plaster. 



Solution 
From Table 10-4, 




8 in. cinder 




aggregate block 


C=0.60 


Corkboard 


k = 0.30 


Cement plaster 


k =8.00 


From Table 




10-5,4, 




inside 




surface 




conductance 


ft = 1-65 


outside 




surface 




conductance 


/„ = 4.00 


Applying 




Equation 10-5, 


-I+-L+-L 


the over-all 


4 0.6 0.3 


thermal resist- 


. 0.5 1 


ance, 1/17 


+ T + L65 




= 0.25 + 1.667 + 13.333 




+ 0.0625 + 0.607 




= 15.92 


Therefore, U 


= 1/15.92 




= 0.0622 Btu/hr/sq ft/ F 



For the most part, it is the insulating material 
used in the wall construction that determines 
the value of U for cold storage walls. The 
surface conductances and the conductances of 
the other materials in the wall have very little 
effect on the value of U because the thermal 
resistance of the insulating material is so large 
with relation to that of the air films and other 
materials. Therefore, for small coolers, it is 
sufficiently accurate to use the conductance of 
the insulating material alone as the wall U 
factor. 

10-10. Temperature Differential across Cold 
Storage Walls. The temperature differential 
across cold storage walls is usually taken as the 
difference between the inside and outside design 
temperatures. 

The inside design temperature is that which 
is to be maintained inside the refrigerated space 
and usually depends upon the type of product 
to be stored and the length of time the product 
is to be kept in storage. The recommended 
storage temperatures for various products are 
given in Tables 10-10 through 10-13. 



150 PRINCIPLES OF REFRIGERATION 



The outside design temperature depends on 
the location of the cooler. For cold storage 
walls located inside a building, the outside 
design temperature for the cooler wall is taken 
as the inside temperature of the building. When 
cold storage walls are exposed to the outdoors, 
the outdoor design temperature for the region 
(Table 10-6) is used as the outside design tem- 
perature. The outdoor design temperatures 
given in Table 10-6 are average outdoor tem- 
peratures and include an allowance for normal 
variations in the outdoor design dry bulb 
temperature during a 24-hr period. These 
temperatures should not be used for calculating 
air conditioning loads. 

1 0-1 1.TemperatureDifferential across Ceil- 
ings and Floors. When a cooler is located 
inside of a building and there is adequate 
clearance between the top of the cooler and the 
ceiling of the building to allow free circulation of 
air over the top of the cooler, the ceiling of the 
cooler is treated the same as an inside wall. 
Likewise, when the top of the cooler is exposed 
to the outdoors, the ceiling is treated as an 
outdoor wall. The same holds true for floors 
except when the cooler floor is laid directly on a 
slab on the ground. As a general rule, the 
ground temperature under a slab varies only 
slightly the year round and is always consider- 
ably less than the outdoor design dry bulb 
temperature for the region in summer. Ground 
temperatures used in determining the tempera- 
ture differential across the floor of cold storage 
rooms are given in Table 10-6A and are based 
on the regional outdoor design dry bulb 
temperature for winter. 

10-12. Effect of Solar Radiation. Whenever 
the walls of a refrigerator are so situated that 
they receive an excessive amount of heat by 
radiation, either from the sun or from some 
other hot body, the outside surface temperature 
of the wall will usually be considerably above 
the temperature of the ambient air. A familiar 
example of this phenomenon is the excessive 
surface temperature of an automobile parked 
in the sun. The temperature of the metal 
surface is much higher than that of the sur- 
rounding air. The amount by which the surface 
temperature exceeds the surrounding air tem- 
perature depends upon the amount of radiant 
energy striking the surface and upon the 
reflectivity of the surface. Recall (Section 2-21) 



that radiant energy waves are either reflected 
by or absorbed by any opaque material that 
they strike. Light-colored, smooth surfaces 
will tend to reflect mbre and absorb less radiant 
energy than dark, rough-textured surfaces. 
Hence, the surface temperature of smooth, 
light-colored walls will be somewhat lower than 
that of dark, rough-textured walls under the 
same conditions of solar radiation. 

Since any increase in the outside surface 
temperature will increase the temperature 
differential across the wall, the temperature 
differential across sunlit walls must be corrected 
to compensate for the sun effect. Correction 
factors for sunlit walls are given in Table 10-7. 
These values are added to the normal tempera- 
ture differential. For walls facing at angles to 
the directions listed in Table 10-7, average 
values can be used. 

10-13. Calculating the Wall Gain Load. In 
determining the wall gain load, the heat gain 
through all the walls, including the floor and 
ceiling, must be taken into account. When 
the several walls or parts of walls are of different 
construction and have different U factors, the 
heat leakage through the different parts is 
computed separately. Walls having identical 
U factors may be considered together, provided 
that the temperature differential across the 
walls is the same. Too, where the difference 
in the value of U is slight and/or the wall area 
involved is small, the difference in the U factor 
can be ignored and the walls or parts of walls 
can be grouped together for computation. 

Example 10-5. A walk-in cooler, 16 ft x 
20 ft x 10 ft high is located in the southwest 
corner of a store building in Dallas, Texas (Fig. 
10-2). The south and west walls of the cooler 
are adjacent to and a part of the south and west 
walls of the store building. The store has a 14 ft 
ceiling so that there is a 4 ft clearance between 
the top of the cooler and the ceiling of the store. 
The store is air conditioned and the temperature 
inside the store is maintained at approximately 
80° F. The inside design temperature for the 
cooler is 35° F. Determine the wall gain load 
for the cooler if the walls of the cooler are of the 
following construction: 

South and west 
(outside walls) 6 in. clay tile 
6 in. corkboard 
0.S cement plaster finish 
on inside 



COOLING LOAD CALCULATIONS 



151 



North and east 
(inside walls) 



Ceiling 



Floor 



Solution 

Wall surface area 

North wall 

West wall 

South wall 

East wall 

Ceiling 

Floor 



1 in. board on both sides 

of 2 x 4 studs 
3f in. granulated cork 
Same as north and east 

walls 
4 in. corkboard laid on 

5 in. slab and finished 

with 3 in. of concrete 



10 x 16 = 160 sq ft 
10 x 20 = 200 sq ft 
10 x 16 = 160sqft 
10 x 20 = 200 sq ft 
16 x 20 = 320 sq ft 
16 x 20 = 320 sq ft 
Wall U factors (Tables 10-1, 10-2, and 10-3) 



North and east 

walls 
South and west 

walls 
Ceiling 
Floor 
From Table 10-6, 
outside summer de- 
sign dry bulb for 
Dallas 

From Table 
10-6A, design 
ground temperature 
for Dallas 



0.079 Btu/hr/sq ft/° F 



0.045 
0.079 
0.066 



92° F 



70° F 



20' : 



, Outside design 

16 ' H temperature, 92* F 



M I MM I MIM ■ ■ 



6 in. corkboard-' | 6 in. clay tile 

Cooler 35° F 
10 ft ceiling 



Partitions 3| in. 

granulated cork- 
1 in. board 
on each side 



^ mm 



Inside temperature 

80' F 

Ceiling 14 ft 



Fig. 10-2 



A short method calculation may be used to 
determine the wall gain load for small coolers 
and for large coolers where the wall U factor 



Outside 
Design 
Temp. 



Inside 
Design 
Temp. 



Normal 
Wall 
T.D. 



Correction 

Factor from 

Table 10-7 



Design 
Wall 
T.D. 



North wall 
South wall 
West wall 
East wall 
Ceiling 
Floor 



80° F 
92° F 
92° F 
80° F 
80° F 
70° F 



35° F 
35° F 
35* F 
35° F 
35° F 
35° F 



45° F 
57° F 
57= p 
45° F 
45° F 
35° F 





4°F 

6°F 









45° F 
61° F 
63° F 
45° F 
45° F 
35° F 



Applying Equation 10-2, 

North wall 160 x 0.079 x 45 = 569Btu/hr 

West wall 200 x 0.045 x 63 = 567 

South wall 160 x 0.045 x 61 = 439 

East wall 200 x 0.079 x 45 = 7ll 

Ceiling 320 x 0.079 x 45 = 1,137 

Floor 320 x 0.066 x 35 = 739 

Total wall gain load 4 . 162 Btu/hr 

= 4,162 x 24 - 99,890 Btu/24 hr 



and temperature difference are approximately 
the same for all the walls. Table 10-18 lists wall 
gain factors (Btu/24 hr sq ft) based on the thick- 
ness of the wall insulation and on the tempera- 
ture differential across the wall. To compute 
the wall gain load in Btu/24 hr by the short 
method, multiply the total outside wall area 
(including floor and ceiling) by the appropriate 
wall gain factor from Table 10-18, viz: 



152 PRINCIPLES OF REFRIGERATION 



Wall gain load = Outside surface area 

x wall gain factor 

To select the appropriate wall gain factor 
from Table 10-18, find the thickness of the wall 
insulation in the extreme left-hand column of 
the table, move right to the column headed by 
the design wall temperature difference, and read 
the wall gain factor in Btu/24 hr/sq ft. For 
example, assume that the walls of a cooler are 
insulated with the equivalent of 4 in. of cork- 
board and that the temperature difference 
across the walls is 55° F. From Table 10-18, 
read the wall gain factor of 99 Btu/24 hr/sq ft 
(see Example 10-18). 

10-14. Calculation the Air Change Load. 
The space heat gain resulting from air changes 
in the refrigerated space is difficult to determine 
with any real accuracy except in those few cases 
where a known quantity of air is introduced 
into the space for ventilating purposes. When 
the weight of outside air entering the space in a 
24-hr period is known, the space heat gain 
resulting from air changes depends upon the 
difference in the enthalpy of the air at the 
inside and outside conditions and can be 
calculated by applying the following equation: 

Air change load = W(h„ - hi) (10-6) 

where W = weight of air entering space in 24 hr 
(lb/24 hr) 
h, = enthalpy of outside air (Btu/lb) 
hi = enthalpy of inside air (Btu/lb) 

However, since air quantities are usually 
given in cubic feet rather than in pounds, to 
facilitate calculations the heat gain per cubic 
foot of outside air entering the space is listed in 
Tables 10-8A and 10-8B for various inside and 
outside air conditions. To determine the air 
change load in Btu per 24 hr, multiply the air 
quantity in cubic feet per 24 hr by the appropri- 
ate factor from Table 10-8 A or 10-8B. 

Where the ventilating air (air change) quantity 
is given in cubic feet per minute (cfm), convert 
cfm to cubic feet per 24 hr by multiplying by 
60 min and by 24 hr. 

Example 10-6. Three hundred cfm of air 
are introduced into a refrigerated space for 
ventilation. If the inside of the cooler is main- 
tained at 35° F and the outside dry bulb tem- 
perature and humidity are 85° F and 50%, 
respectively, determine the air change load in 
Btu/24 hr. 



Solution 
Cubic feet of air per 
per24hr 



From Table 10-8.A 
heat gain per cubic 
feet 

Ventilating (air 
change) load 



cfm x 60 x 24 
300 x 60 x 24 
432,000 cu ft/24 hr 



= 1.86Btu/cuft 

= cu ft/24 hr 
x Btu/cu ft 
= 432,000 x 1.86 
= 803,520 Btu/24 hr 



Except in those few cases where air is pur- 
posely introduced into the refrigerated space for 
ventilation, the air changes occurring in the 
space are brought about solely by infiltration 
through door openings, The quantity of outside 
air entering a space through door openings in 
a 24-hr period depends upon the number, size, 
and location of the door or doors, and upon 
the frequency and duration of the door open- 
ings. Since the combined effect of all these 
factors varies with the individual installation 
and is difficult to predict with reasonable ac- 
curacy, it is general practice to estimate the 
air change quantity on the basis of experience 
with similar applications. Experience has shown 
that, as a general rule, the frequency and dur- 
ation of door openings and, hence, the air 
change quantity, depend on the inside volume 
of the cooler and the type of usage. Tables 
10-9A and 10-9B list the approximate number 
of air changes per 24 hr for various cooler 
sizes. The values given are for average usage 
(see table footnotes). The ASRE Data Book 
defines average and heavy usage as follows: 

Average usage includes installations not 
subject to extreme temperatures and where 
the quantity of food handled in the refrigerator 
is not abnormal. Refrigerators in delicatessens 
and clubs may generally be classified under this 
type of usage. 

Heavy usage includes installations such as 
those in busy markets, restaurant and hotel 
kitchens where the room temperatures are likely 
to be high, where rush periods place heavy loads 
on the refrigerator, and where large quantities 
of warm foods are often placed in it.* 

* The Refrigerating Data Book, Basic Volume, 
The American Society of Refrigerating Engineers, 
1949, New York, p. 327. 



Example 10-7. A walk-in cooler 8 ft x 
IS ft x 10 ft high is constructed of 4 in. of cork- 
board with 1 in. of wood on each side. The 
outside temperature is 95° F and the humidity 
is 50%. The cooler is maintained at 35° F and 
the usage is average. Determine the air change 
loadinBtu/24hr. 

Solution. Since the walls of the cooler are 
approximately 6 in. thick (4 in. of corkboard 
and 2 in. of wood), the inside dimensions of the 
cooler are 1 ft less than the outside dimensions; 
therefore, 
Inside volume = 7 ft x 14 ft x 9 ft 

= 882 cu ft 
From Table 10-9A, 
by interpolation, num- 
ber of air changes per 
24 hr for cooler vol- 
ume of approximately 
900cuft =19 

Total quantity of 
air change per 24 hr = Inside volume 

x air changes 
= 882 x 19 
= 16,758 cu ft/24 hr 
From Table 10-8A, 
heat gain per cubic 
feet = 2.49 Btu/cu ft 

Air change load = cu ft/24 hr 

x Btu/cu ft 
= 16,758 x 2.49 
= 41,727 Btu/24 hr 

10-15. Calculation the Product Load. When 
a product enters a storage space at a tempera- 
ture above the temperature of the space, the 
product will give off heat to the space until it 
cools to the space temperature. When the 
temperature of the storage space is maintained 
above the freezing temperature of the product, 
the amount of heat given off by the product in 
cooling to the space temperature depends Upon 
the temperature of the space and upon the 
weight, specific heat, and entering temperature 
of the product. In such cases, the space heat 
gain from the product is computed by the 
following equation, (see Section 2-24): 

Q = W x C x(T a -Tj (10-7) 
where Q = the quantity of heat in Btu 

W = weight of the product (pounds) 
C = the specific heat above freezing 

(Btu/lb/° F) 
T x = the entering temperature (° F) 
T a = the space temperature (° F) 



COOLING LOAD CALCULATIONS 153 

Example 10-8. One thousand pounds of 
fresh, lean beef enter a cooler at 55° F and are 
chilled to the cooler temperature of 35° F in 
24 hr. Calculate the product load in Btu/24 hr. 

Solution 
From Table 10-12, the 
specific heat of lean 
beef above freezing = 0.75 Btu/lb/° F 

Applying Equation 
10-7, the product load, 
Btu/24 hr = 1000 x 0.75 

x (55 - 35) 
= 1000 x 0.75 x 20 
= 15,000 Btu/24 hr 

Notice that no time element is inherent in 
Equation 10-7 and that the result obtained is 
merely the quantity of heat the product will 
give off in cooling to the space temperature. 
However, since in Example 10-8 the product is 
to be cooled over a 24-hr period, the resulting 
heat quantity represents the product load for a 
24-hr period. When the desired cooling time 
is less than 24 hr, the equivalent product load 
for a 24-hr period is computed by dividing the 
heat quantity by the desired cooling time for the 
product to obtain the hourly cooling rate and 
then multiplying the result by 24 hr to determine 
the equivalent product load for a 24-hr period. 
When adjusted to include these two factors, 
Equation 10-7 is written: 

WxCx(r,-r 1 )x24hr 

Q aav J n*s 0°- 8 ) 

desired cooling time (hr) 

Example 104. Assume that it is desired to 
chill the beef in Example 10-8 in 6 hr rather than 
in 24 hr. Determine the product load in 
Btu/24 hr. 



Solution. Applying 
Equation 10-8, product 
load, Btu/24 hr 



1000 x 0.75 
_ x (55 - 35) x 24 

6 
= 60,000 Btu/24 hr 



Compare this result with that obtained in 
Example 10-8. 

When a product is chilled and stored below 
its freezing temperature, the product load is 
calculated in three parts: 

1. The heat given off by the product in cooling 
from the entering temperature to its freezing 
temperature. 



154 PRINCIPLES OF REFRIGERATION 



2. The heat given off by the product in solidi- 
fying or freezing. 

3. The heat given off by the product in cool- 
ing from its freezing temperature to the final 
storage temperature. 

For parts 1 and 3, Equation 10-7 is used to 
determine the heat quantity. For part 1, T x in 
Equation 10-7 is the entering temperature of the 
product, whereas T z is the freezing temperature 
of the product (Tables 10-10 through 10-13). 
For part 3, T x in Equation 10-7 is the freezing 
temperature of the product and T t is the final 
storage temperature. The heat quantity for 
part 2 is determined by the following equation: 

Q = W x h if (10-9) 

where W = the weight of the product (pounds) 
h it = the latent heat of the product 
(Btu/lb) 

When the chilling and freezing are accom- 
plished over a 24-hr period, the summation of 
the three parts represents the product load for 
24 hr. When the desired chilling and freezing 
time for the product are less than 24 hr, the 
summation of the three parts is divided by the 
desired processing time and then multiplied by 
24 hr to determine the equivalent 24-hr product 
load. 

Example 10-10. 500 pounds of poultry 
enter a chiller at 40° F and are frozen and chilled 
to a final temperature of — 5° F for storage in 
12 hr. Compute the product load in Btu/24 hr. 

Solution 
From Table 10-12, 
Specific heat above 

freezing 
Specific heat below 

freezing 
Latent heat 
Freezing 

temperature 

To cool poultry from 
entering temperature 
to freezing tempera- 
ture, applying Equa- 
tion 10-7 



= 0.79 Btu/lb/° F 

= 0.37 Btu/lb/° F 
= 106 Btu/lb 

= 27°F 



To freeze, applying 
Equation 10-9 



500 x 0.79 
x (40 - 27) 
5135 Btu 

500 x 106 
53,000 Btu 



To cool from freezing 
temperature to final 
storage temperature, 
applying Equation 
10-7 



500 x 0.37 
x[27-(-5)] 

5920 Btu 



Total heat given up 
by product (summa- 
tion of 1,2, and 3) 

Equivalent product 
load for 24-hrperiod 
Btu/24 hr 



= 64,000 Btu 

64,000 x 24 hr 

12 hr 

= 128,000 Btu/24 hr 

10-16. Chilling Rate Factor. During the early 
part of the chilling period, the Btu per hour load 
on the equipment is considerably greater than the 
average hourly product load as calculated in 
the previous examples. Because of the high 
temperature difference which exists - between the 
product and the space air at the start of the 
chilling period, the chilling rate is higher and 
the product load tends to concentrate in the 
early part of the chilling period (Section 9-23). 
Therefore, where the equipment selection is 
based on the assumption that the product load 
is evenly distributed over the entire chilling 
period, me equipment selected will usually have 
insufficient capacity to carry the load during 
the initial stages of chilling when the product 
load is at a peak. 

To compensate for the uneven distribution of 
the chilling load, a chilling rate factor is intro- 
duced into the chilling load calculation. The 
effect of the chilling rate factor is to increase the 
product load calculation by an amount sufficient 
to make the average hourly cooling rate approxi- 
mately equal to the hourly load at the peak 
condition. This results in the selection of 
larger equipment, having sufficient capacity to 
carry the load during the initial stages of chilling. 

Chilling rate factors for various products are 
listed in Tables 10-10 through 10-13. The 
factors given in the tables are based on actual 
tests and on calculations and will vary with the 
ratio of the loading time to total chilling time. 
As an example, test results show that in typical 
beef and hog chilling operations the chilling rate 
is 50% greater during the first half of the chill- 
ing period than the average chilling rate for 
the entire period. The calculation without the 
chilling rate factor will, of course, show the 



average chilling rate for the entire period. To 
obtain this rate during the initial chilling period, 
it muSt be multiplied by l.S. For convenience, 
the chilling rate factors are given in the tables 
in reciprocal form and are used in the denomi- 
nator of the equation. Thus the chilling rate 
factor for beef as shown in the table is 0.67 
0/1.5). 

Where a chilling rate factor is used, Equation 
10-7 is written 

W xC xjTi-TJ 
Q ~ Chilling rate factor (1 °" 10) 

As a general rule chilling rate factors are not 
used for the final stages of chilling from the 
freezing temperature to the final storage tem- 
perature of the product. Too, chilling rate 
factors are usually applied to chilling rooms 
only and are not normally used in calculation 
of the product load for storage rooms. Since 
the product load for storage rooms usually 
represents only a small percentage of the total 
load, the uneven distribution of the product 
load over the cooling period will not ordinarily 
cause overloading of the equipment and, there- 
fore, no allowance need be made for this 
condition. 

10-17. Respiration Heat. Fruits and vegetables 
are still alive after harvesting and continue to 
undergo changes while in storage. The more 
important of these changes are produced by 
respiration, a process during which oxygen 
from the air combines with the carbohydrates 
in the plant tissue and results in the release of 
carbon dioxide and heat. The heat released is 
called respiration heat and must be considered 
as a part of the product load where considerable 
quantities of fruit and/or vegetables are held in 
storage. The amount of heat evolving from the 
respiration process depends upon the type and 
temperature of the product. Respiration heat 
for various fruits and vegetables is listed in 
Table 10-14. 

Since respiration heat is given in Btu per 
pound per hr, the product load accruing from 
respiration heat is computed by multiplying the 
total weight of the product by the respiration 
heat as given in Table 10-14, viz: 

Q (Btu/24 hr) - Weight of product (lb) 

x respiration heat (Btu/lb/hr) 
x24hr (10-11) 



COOLING LOAD CALCULATIONS 155 

10-18. Containers and Packing Materials. 

When a product is chilled in containers, such as 
milk in bottles or cartons, eggs in crates, fruit 
and vegetables in baskets and lugs, etc., the 
heat given off by the containers and packing 
materials in cooling from the entering tempera- 
ture to the 'space temperature must be con- 
sidered as a part of the product load. Equation 
10-7 is used to compute this heat quantity. 
10-19. Calculating the Miscellaneous Load. 
the miscellaneous load consists primarily of the 
heat given off by lights and electric motors 
operating in the space and by people working 
in the space. The heat given off by lights is 
3.42 Btu per watt per hour. The heat given off 
by electric motors and by people working in the 
space is listed in Tables 10-15 and 10-16, 
respectively. The following calculations are 
made to determine the heat gain from mis- 
cellaneous: 

Lights: wattage x 3.42 Btu/watt/hr x 24 hr 

Electric motors : factor (Table 10-1 5) x horse- 
power x 24 hr 

People: factor (Table 10-16) x number of 
people x 24 hr 

10-20. Use of Safety Factor. The total cooling 
load for a 24-hr period is the summation of the 
heat gains as calculated in the foregoing sections. 
It is common practice to add 5% to 10% to 
this value as a safety factor. The percentage 
used depends upon the reliability of the informa- 
tion used in calculating the cooling load. As a 
general rule 10% is used. 

After the safety factor has been added, the 
24-hr load is divided by the desired operating 
time for the equipment to determine the average 
load in Btu per hour (see Section 10-2). The 
average hourly load is used as a basis for 
equipment selection. 

10-21. Short Method Load Calculations. 
Whenever possible the cooling load should be 
determined by using the procedures set forth 
in the preceding sections of this chapter. How- 
ever, when small coolers (under 1600 cu ft) are 
used for general-purpose storage, the product 
load is frequently unknown and/or varies 
somewhat from day to day so that it is not pos- 
sible to compute the product load with any real 
accuracy. In such cases, a short method of load 
calculation can be employed which involves the 
use of load factors which have been determined 
by experience. When the short method of 



156 PRINCIPLES OF REFRIGERATION 



calculation is employed, the entire cooling load 
is divided into two parts: (1) the wall gain load 
and (2) the usage or service load. 

The wall gain load is calculated as outlined 
in Section 10-13. The usage load is computed 
by the following equation: 

Usage load = interior volume x usage factor 

(10-12) 

Notice that the usage factors listed in Table 
10-17 vary with the interior volume of the cooler 
and with the difference in temperature between 
the inside and outside of the cooler. Too, an 
allowance is made for normal and heavy usage. 
Normal and heavy usage have already been 
defined in Section 10-14. No safety factor is 
used when the cooling load is calculated by 
the short method. The total cooling load is 
divided by the desired operating time for the 
equipment to find the average hourly load used 
to select the equipment. 

Example 10-11. A walk-in cooler 18 ft x 
10 ft x 10 ft high has 4 in. of corkboard insula- 
tion and standard wall construction consisting 
of two layers of paper and 1 in. of wood on each 
side (total wall thickness is 6 in.). The tempera- 
ture outside the cooler is 85° F. 35001b of 
mixed vegetables are cooled 10° F to the storage 
temperature each day. Compute the cooling 
load in Btu/hr based on a 16-hr per day opera- 
ting time for the equipment. The inside 
temperature is 40° F. 

Solution 
Outside surface area = 2 x 18 ft x 10 ft 

- 360 sq ft 
= 4 x 10 ft x 10 ft 
= 400 sq ft 
760 sq ft 
Inside volume (since 
total wall thickness is 
6 in., the inside dimen- 
sions are 1 ft less than 
the outside dimensions) 



Wall gain factor 
(Table 10-18) (45° F 
TD and 4 in. insulation) 

Air changes (Table 
10-9A) (by interpola- 
tion) 

Heat gain per cubic 
foot of air (Table 
10-8A)(50%RH) 



17ft x9ft x9ft 
1377 cu ft 



81 Btu/hr/sq ft 
■■ 16.8 per 24 hr 
• 1.69 Btu/cuft 



Average specific heat 
of vegetables (Table 
10-11) = 0.9 Btu/lb/° F 

Average respiration 
heat of vegetables 
(Table 10-14) = 0.09 Btu/lb/hr 

Wall gain load: 
Area x wall gain factor 
= 760 sq ft x 81 Btu/sq ft/24 hr 

= 61,560 Btu/24 hr 
Air change load: 
Inside volume 

x air changes x Btu/cuft 
= 1377cuft x 16.8 
x 1.69 Btu/cuft = 23,100 Btu/24 hr 
Product load: 
Temperature reduction 

= Mxcx(r,-r 1 ) 

= 35001b x0.9Btu/lb/°F 

x 10° F = 31,500 Btu/24 hr 

Respiration heat 
= M x reaction heat x 24 hr 
= 3500 x 0.09 Btu/lb/hr 
x24hr 

7,560 Btu/24 hr 
Summation: 
Safety factor 

(10%) =1 2,370 Btu 

Total cooling 

load = 136,100 Btu/24 hr 

Required 
cooling 
capacity 
(Btu/hr) _ Total cooling load 



123,720 Btu/24 hr 



Desired running time 
_ 136,100 Btu/24 hr 

16 hr 
= 8,500 Btu/hr 

Example 10-12. The dimensions of a 
banana storage room located in New Orleans, 
Louisiana are 22 ft x 32 ft x 9 ft. The walls 
are 1 in. boards on both sides of 2 x 4 studs and 
insulated with 3$ in. granulated cork. The floor 
and roof are of the same construction as the 
walls. The floor is over a ventilated crawl space 
and the roof is exposed to the sun. The tempera- 
ture outside the storage room is approximately 
the same as the outdoor design temperature for 
the region. 30,000 lb of bananas are ripened at 
70° F and then cooled to 56° F in 12 hr for 
holding storage. Compute the Btu/hr cooling 
load. (Note: Because of the high storage tem- 
perature the evaporator will not collect frost and 
the equipment is designed for continuous run.) 



COOLING LOAD CALCULATIONS 157 



= 704 sq ft 
= 1676sqft 

- 5208 cu ft 



=*89°F 

= 0.079Btu/hr/sqft/°F 

= 20°F 



= 7/24hr 



Solution 
Outside surface area 
Ceiling 
Walls and floor 

Inside volume 
(21 ft x 31 ft x 8 ft) 

Outside design 
temperature for 
New Orleans (Table 
10-6) 

Wall U factor 
(Table 10-2) 

Sun factor for roof 
(Table 10-7) (tar 
roofing) 

Air changes (Table 
10-9 A) (by interpo- 
lation) 

Heat gain per 
cubic foot of air 

change(Table 10-8A) = 1 .4 Btu/cu ft 
(interpolated) 

Specific heat of 
bananas (Table 
10-10) 

Reaction heat of 
bananas (Table 
10-14) 
Wall gain load: 

Area x U x TD x 24 hr 
Ceiling 
704 x 0.079 x 53 x 24 

= 70,740 Btu/24 hr 
Walls and floor 

1676 x 0.79 x 33 x 24 

= 104,860 Btu/24 hr 
Air change load: 
Inside volume 

x air changes x Btu/cu ft 
= 5208 x 7 x 1.4 Btu/cu ft 

= 51,000 Btu/24 hr 
Product load: 
Temperature reduction 
_ M x C xjTj-Tj) x24hr 

Chilling time (hr) 
_ 30,000 x 0.9 x 14 x 24 
12 

= 756,000 Btu/24 hr 
Respiration heat 
= M x reaction heat x 24 hr 
= 30,000 x 0.5 x 24 = 360,000 Btu/24 hr 
Summation: 



= 0.9Btu/lb/°F 
= 0.5 Btu/lb/hr 



Safety factor (10%) 
Total cooling load 



1,342,600 Btu/24 hr 

1 34,260 Btu 
1,476,860 Btu/24 hr 



Required cooling capacity 
(Btu/hr) 

Total cooling load 
Desired running time 

_ 1,476,860 Btu/24 hr 
24 hr 

= 61,530 Btu/hr 

Example 10-13. A chilling room 35 ft x 
50 ft x 15 ft is used to chill 50,000 lb of fresh 
beef per day from an initial temperature of 
100° F to a final temperature of 35° F in 18 hr. 
Four people work in the chiller during the 
loading period. The lighting load is 1500 watts. 
The floor, located over an unconditioned space, 
is a 5 in. concrete slab insulated with 4 in. of 
corkboard and finished with 3 in. of concrete. 
The ceiling, situated beneath an unconditioned 
space, is a 4 in. concrete slab with wood sleepers 
and insulated with 4 in. of corkboard. All of 
the walls are inside partitions adjacent to uncon- 
ditioned spaces (90° F) except the east wall 
which is adjacent to a 35° F storage cooler. Wall 
construction is 4 in. cinder block insulated with 
4 in. of corkboard and finished pn one side with 
plaster. Compute the cooling load in Btu per 
hour based on a 16-hr operating time for the 
equipment. 

Solution 

Outside surface area 
Ceiling 
(35 ft x50ft) 
Floor 

(35 ft x50ft) 
Walls (except 
east) 

(120 ft x 15 ft) 
Inside volume 
(34 ft x 49 ft x 
13.5 ft) 

Air changes 
(Table 10-9A) (by 
interpolation) 

Heat gain per 
cubic foot (Table 
10-8A)(50%RH) 

Specific heat of 
beef (Table 10-12) 
Chilling rate factor 
(Table 10-12) 
Occupancy heat 
gain (Table 10-16) 
Ceiling U factor 
(Table 10-3) 



= 1,750 sq ft 
= 1,750 sq ft 

= 1,800 sq ft 

- 22,491 cu ft 

= 3.2 per 24 hr 

= 2.17 Btu/cu ft 
= 0.75 Btu/lb/° F 
= 0.67 

= 900 Btu/hr/person 
= 0.069 Btu/hr/sqft/°F 



158 PRINCIPLES OF REFRIGERATION 



Floor U factor 

(Table 10-3) = 0.066 Btu/hr/sq ft/°F 

Wall U factor 

(Table 10-1) = 0.065 Btu/hr/sq ft/°F 

Wall gain load: 
A x U xTD x 24 hr 
= 5300 x 0.067 x55 x24hr 

= 468,700 Btu/24 hr 
Air change load: 
Inside volume 

x air changes x Btu/cu ft 
= 22,491 x 3.2 x 2.17 

= 156,000 Btu/24 hr 
Product load: 

M x C x (T % - T x ) x 24 
Chilling time (hr) x chilling rate factor 
50,000 x 0.75 x 65 x 24 
18 x 0.67 

= 4,850,700 Btu/24 hr 
Miscellaneous: 
Occupancy 

= No. of people x factor x 24 hr 
= 4 x900 x24 

= 86,400 Btu/24 hr 
Lighting load 
= watts x 3.4 x 24 hr 
1500 x 3.4 x 24 = 122,400 Btu/24 hr 
Summation: 5,684,200 Btu/24 hr 

Safety factor (10%) 

= 568,420 Btu 
6,252,620 Btu/24 hr 
Required cooling capacity (Btu/hr) 

_ Total cooling load 
Desired operating time 
6,252,620 Btu/24 hr 
18 hr 
= 390,800 Btu/hr 

Note: (1) Since there is no temperature 
differential across the east wall, there is no gain 
or loss of heat through the wall and the wall 
is ignored in the cooling load calculations 
However, this wall should be insulated with at 
least the minimum amount of insulation to 
prevent excessive heat gains through this wall 
in the event that the adjacent refrigerated space 
should become inoperative. (2) Since the TD 
across all the walls, including floor and ceiling, 
is the same and since the difference in the wall 
U factors is slight, the walls may be lumped 
together for calculation. (3) Although the 
workmen are in the space for only 4 hr each 
day for the purpose of load calculation, they are 
assumed to be in the cooler continuously. This 
is because their occupancy occurs simultan- 



eously with the chilling peak. If the occupancy 
occurred at any time other than at the peak, 
the occupancy load could be ignored. 

Example 10-14. Three thousand lug boxes 
of apples are stored at 35° F in a storage cooler 
50 ft x 40 ft x 10 ft. The apples enter the cooler 
at a temperature of 90° F and at the rate of 200 
lugs per day each day for the 15 day harvesting 
period. The walls including floor and ceiling 
are constructed of 1 in. boards on both sides of 
2x4 studs and are insulated with 3$ in. of 
rock wool. All of the walls are shaded and the 
ambient temperature is 85° F. The average 
weight of apples per lug box is 59 lb. The lug 
boxes have an average weight of 4.5 lb and a 
specific heat value of 0.60 Btu/lb/° F. Determine 
the average hourly cooling load based on 16 hr 
operating time for the equipment. 

Solution 
Outside surface 
area = 5800 sq ft 

Inside volume 
(49 ft x 39 ft x 9 ft) = 17,200 cu ft 

Wall U factor 
(Table 10-2) = 0.072 Btu/hr/sq ft/ F 

Air changes 
(Table 10-9A) (by 
interpolation) = 3.7 per 24 hr 

Heat gain per 
cubic foot (Table 
10-8 A) = 1.86 Btu/cu ft 

Specific heat of 
apples (Table 10-10) = 0.89 Btu/lb/° F 

Respiration heat 
of apples (Table 
10-14) (by interpo- 
lation) = 0.025 Btu/lb/hr 

Wall gain load: 
A x U x TD x 24 hr 
= 5800 x 0.072 x 50 x 24 

= 501,100 Btu/24 hr 
Air change load: 
Inside volume x air changes 

x Btu/cu ft 
= 17,000 x 3.7 x 1.86 

= 11 7,000 Btu/24 hr 
Product load: 
Temperature reduction 
_ M x C x (T t - T x ) 
Chilling rate factor 
Apples 
(200 x 59 lb) x 0.89 x 55 



0.67 



= 862,100 Btu/24 hr 



COOLING LOAD CALCULATIONS 159 



Fig. 103 




Lug boxes 
(200 x 4.5 lb) x 0.6 x 55 
0.67 

= 44,300 Btu/24 hr 
Respiration 

= M x reaction heat x 24 hr 
= (3000 x 59 lb) x 0.025 x 24 

= 106,200 Btu/24 hr 



Summation 

Safety factor (10%) 
Total cooling load 

Average hourly load = 



1,630,700 Btu/24 hr 
= 163,100 Btu 
= 1,793,800 Btu/24 hr 
_ Total cooling load 

Running time 
_ 1,793,800 Btu/24 hr 

16 hr 
= 112,100 Btu/24 hr 

Note: Load calculation and equipment 
selection is based on maximum loading which 
occurs on the fifteenth day. 

Example 10-15. Twenty-two thousand 
pounds of dressed poultry are blast frozen on 
hand trucks each day (24 hr) in a freezing tunnel 
14 ft x 9 ft x 10 ft high (see Fig. 10-3). The 
poultry is precooled to 45° F before entering the 
freezer where it is frozen and its temperature 
lowered to 0° F for storage. The lighting load 
is 200 watts. The hand trucks carrying the 
poultry total 1400 lb per day and have a specific 
heat of 0.25 Btu/lb/° F. The partitions adjacent 
to the equipment room and vestibule are con- 
structed of 6 in. clay tile insulated with 8 in. of 
corkboard. Partitions adjacent to storage 
cooler are 4 in. clay tile with 2 in. corkboard 
insulation. The roof is a 6 in. concrete slab 
insulated with 8 in. of corkboard and covered 
with tar, felt, and gravel. The floor is a 6 in. 
concrete slab insulated with 8 in. of corkboard 



and finished with 4 in. of concrete. The floor is 
over a ventilated crawl space. Roof is exposed 
to the sun. The equipment room is well venti- 
lated so that the temperature inside is approxi- 
mately the outdoor design temperature for the 
region. The storage room is maintained at 0° F, 
whereas the temperature in the freezer is — 10° F. 
The location is Houston, Texas. Determine the 
average hourly refrigeration load based on 20 hr 
per day operating time for the equipment. 

Solution 

Outside surface area 
Roof 

(9 ft + 14 ft) =126sqft 

Floor 

(9 ft x 14 ft) =126sqft 

N and E partitions 
(23 ft x 10 ft) =230sqft 
S and W partitions 
(23 ft x 10 ft) =230sqft 
Inside volume 
(8ft x9ft x 13ft) 
Summer outdoor 
design temperature 
U factors 
Roof 

(Table 10-3) 
Floor 
(Table 10-3) 



= 936 cu ft 



92° F 



= 0.036 Btu/hr/sq ft/° F 



= 0.035 Btu/hr/sq ft/° F 
N and E partitions 

(Table 10-2) = 0.035 Btu/hr/sq ft/° F 
S and W partitions 

(Table 10-2) =0.12 Btu/hr/sq ft/° F 



Roof sun factor 
(Table 10-7) 

Air changes 
(Table 10-9B) 



= 20°F 



13.5 per 24 hr 



Heat gain per cubic foot 
(Table 10-8B) = 3.56 Btu/cu ft 



160 PRINCIPLES OF REFRIGERATION 



Cold 
18' storage 
38*F 



— »-N 
-32.5*- 



Lochtrs 



O'F 



Lockers 



Lockers 



Freezer 



> 



Lockers 



N- 



Cold storage 
38"F 



Conditioned 
space, 80* F 



Fig. 10-4. Frozen food locker plant. 

Specific heat of poultry 
(Table 10-12) 

Above freezing 

= 0.79 Btu/lb/° F 
Below freezing 

= 0.37 Btu/lb/° F 
Latent heat of poultry 
(Table 10-12) 

= 106 Btu/lb 
Freezing temperature 

27° F 
Wall gain load: 
^xt7xri)x24hrs 
Floor 

126 x 0.035 x 102 x 24 

= 10,800 Btu/24 hr 
Roof 
126 x 0.036 x (102 + 20) x 24 

= 13,300 Btu/24 hr 
South and west partition 
230 x 0.035 x 102 x 24 

= 19,700 Btu/24 hr 
North and east partition 
230 x 0.12 x 10 x 24 

= 6,600 Btu/24 hr 
Air change load: 
Inside volume x air changes 
x Btu/cu ft 
- 936 x 13.5 x 3.56 

= 45,000 Btu/24 hr 
Product Load: 
Temperature reduction 
= M x C x(r s -TJ 
Poultry 

_ 22,000 x 0.79 x (45 - 27) 
0.67 (chilling factor) 

= 302,700 Btu/24 hr 
22,000 x 0.37 x (27 - 0) 

- 219,700 Btu/24 hr 
Trucks 

1,400 x 0.25 x (92 - 0) 



Freezing = M x latent heat 
Poultry = 22,000 x 106 

= 2,332,000 Btu/24 hr 
Miscellaneous: 
Lighting: 200 watts x 3.4 
Btu/watt/hr x 24 hr 

= 16,300 Btu/24 hr 



Summation: 
Safety factor (10%) 

Total cooling 

load 
Average hourly 

load 



3,014,100 Btu/24 hr 
= 301,400 Btu 

= 3,315,500 Btu/24 hr 

_ 3,315,500 Btu/24 hr 
20 hr (running time) 
= 165,775 Btu/hr 



Example 10-16. A frozen food locker plant 
18 ft x 32.5 ft x 10 ft, containing 353 indi- 
vidual lockers and an 8 ft freezing cabinet, is 
located in Tulsa, Oklahoma (see Fig. 10-4). 
The north and west wall are constructed of 8 in. 
clay tile with 6 in. of corkboard insulation. 
South and east walls are 4 in. clay tile with 4 in. 
of corkboard insulation. The roof is exposed to 
the sun and is constructed of 4 in. of concrete 
insulated with 8 in. of corkboard and covered 
with tar, felt, and gravel. The floor is a 5 in. 
concrete slab poured directly on the ground, 
insulated with 8 in. of corkboard, and finished 
with 3 in. of concrete. The product load on 
cabinet freezer is 700 lb of assorted meats per 
day. (Standard practice is to allow for 2 lb of 
product per locker per day.) In this instance, 
the product is precooled to 38° F before being 
placed in the freezer. The lighting load is 500 
watts and the average occupancy is three people. 
Determine the average hourly refrigerating rate 
based on a 20-hr equipment operating time. 



0.67 



. 48,000 Btu/24 hr 



Solution 

Outside surface area 
Roof (18 

Floor (18 

South and east 
partition (50 

North partition (18 

West wall (32 

Inside volume 
(16 ft x 30.5 ft x 8 



ft x 32.5 ft) 
585 sq ft 
ft x 32.5 ft) 
585 sq ft 

.5 ft x 10 ft) 

505 sq ft 

ft x 10 ft) 

180 sq ft 
.5 ft x 10 ft) 

325 sq ft 

ft) 

3904 cu ft 



COOLING LOAD CALCULATIONS 161 



Design outdoor temperature 
(Table 10-6) = 92°F 

Design ground temperature 
(Table 10-6A) = 65° F 

Roof sun factor 
(Table 10-7) = 20°F 

West wall sun factor 
(Table 10-7) - 6° F 

U factors 
Roof (Table 10-3) 

= 0.036 Btu/hr/sq ft/ F 

Floor (Table 10-3) 

= 0.046 Btu/hr/sq ft/° F 

North and west walls 

(Table 10-1) = 0.034 Btu/hr/sq ft/° F 
South and east walls 

(Table 10-1) = 0.066 Btu/hr/sq ft/° F 

Air changes 
(see Note 2 of Table 10-9B) 

= 12.6(6.3 x 2) 

per 24 hr 
Heat gain per cubic foot 
(Table 10-8B) = 3.0 Btu/cu ft 

Specific heat (average for meat) 
Above freezing = 0.8 Btu/lb/° F 
Below freezing = 0.4 Btu/lb/° F 
Latent heat 

(average) = 100 Btu/lb 

Freezing tempera- 
ture (average) = 28° F 
Occupancy factor 

= 1300 Btu/hr/person 

Wall Gain Load: 
A x U x TD x 24 
Roof 585 x 0.036 

x (92 + 20) x 24 
= 56,600 Btu/24 hr 
Floor 585 x 0.046 x 65 x 24 

= 42,000 Btu/24 hr 
South and east 
partition 505 x 0.066 x 38 x 24 

= 30,400 Btu/24 hr 
North partition 180 x 0.034 x 80 x 24 
= 11,750 Btu/24 hr 

West wall 325 x 0.034 

x (92 + 6) x 24 
= 26,000 Btu/24 hr 

Air change load: 
Inside volume x air changes 
x Btu/cu ft 
= 3904 x 12.6 x 3.01 

= 147,570 Btu/24 hr 



= 531,560 Btu/24 hr 



Product load: 
Temperature reduction 
= M x C x (T t - TJ 
700 x 0.8 x (38 - 28) 

= 5600 Btu/24 hr 
700 x 0.4 x (28 - 0) 

7840 Btu/24 hr 
Freezing 

= M x latent heat 
= 700 x 100 70,000 Btu/24 hr 

Miscellaneous: 
Lights 
= 500 watts x 3.4 Btu/hr 

x 24 hr = 40,800 Btu/24 hr 

Occupancy 
= 3 x 1300 x 24 

= 93,600 Btu/24 hr 

Summation: 
Safety factor 

(10%) = 53,150 Btu 

Total cooling 

load = 584,710 Btu/24 hr 

Average hourly 
load _ 584,710 Btu/24 hr 

20 hr 
= 29,240 Btu/hr 
Load on freezer 91,780 Btu/24 hr 

20 hr 
(product load 
only, including 
10% safety factor) 

= 4,590 Btu/hr 
Load on locker 
room only = 29,240 Btu/hr 

(total load less - 4,590 Btu/hr 

freezer load) = 24,650 Btu/hr 

Example 10-17. Five hundred gallons of 
partially frozen ice cream at 25° F are entering 
a hardening room 10 ft x 15 ft x 9 ft each day. 
Hardening is completed and the temperature of 
the ice cream is lowered to —20° F in 10 hr. 
The walls, including floor and ceiling, are 
insulated with 8 in. of corkboard and the over- 
all thickness of the walls is 12 in. The ambient 
temperature is 90° F and the lighting load is 300 
watts. Assume the average weight of ice cream 
is 5 lb per gallon, the average specific heat below 
freezing is 0.5 Btu/lb/°F, and the average 
latent heat per pound is 100 Btu.* Determine 
the average hourly load based on 18 hr opera- 
tion. 

* These values are variable and depend upon the 
composition of the mix, the percent of overrun, 
and the temperature of the ice cream leaving the 
freezer. 



162 PRINCIPLES OF REFRIGERATION 



Solution 
Outside surface 

area = 750 cu ft 

Inside volume 

(8 ft x 13 ft x 7 ft) 



= 728 cu ft 
99 Btu/sq ft/24 hr 



Wall gain factor 

(Table 10-18) 
Air changes 

(Table 10-9B) 

(interpolated) = 16.7 per 24 hr 

Heat gain per cubic foot 
(Table 10-8B) 
(50 % RH) =3.88 Btu/cu ft 

Wall gain load: 
area x wall gain factor 
= 750 x 99 = 74,250 Btu/24 hr 

Air change load: 
Inside volume x air changes 
x Btu/cu ft 
= 728 x 16.7 x 3.88 

= 47,170 Btu/24 hr 
Product load: 
Temperature reduction 

M x C(T S - 7\) x 24 hr 
— Chilling time 

(500 x 5) x 0.5 x (25 20) x 24 

10 
= 135,000 Btu/24 hr 
Freezing 
_ M x latent heat x 24 

Freezing time 
_ (500 x 5) x 100 x 24 
10 

= 600,000 Btu/24 hr 
Miscellaneous load: 
Lighting: 
300 watts x 3.4 x 24 

= 24,480 Btu/24 hr 

Summation: 880,900 Btu/24 hr 

Safety factor (10%) = 88,090 Btu 

968,990 Btu/24 hr 
Average hourly load 
(968,990/1 8 hr) = 53,800 Btu/hr 

Example 10-18. A cooler 10 ft x 12 ft x 9 
ft is used for general purpose storage in a 
grocery store. The cooler is maintained at 35° F 
and the service load is normal. The walls are 
insulated with the equivalent of 4 in. of cork- 
board and the ambient temperature is 80° F. 
Determine the cooling load in Btu per hour 
based on a 16 hr operating time. 



Solution 
Outside surface area 

Inside volume 
(9 ft x lift x 8 ft) 



= 636 sq ft 



Wall gain factor 

(Table 10-18) 
Usage factor 

(Table 10-17) 

Wall gain load: 
area x wall gain factor 
= 636 x 81 

Usage load: 
Inside volume x 
usage factor 
= 792 x 50 



792 cu ft 

81 Btu/sq ft/24 hr 

50 Btu/cu ft/24 hr 

51,500 Btu/24 hr 



Total cooling load 
Average hourly load 

/ 91,100 Btu/2 4 hr 

I 16 hr 



= 39,600 Btu/24 hr 
= 91,100 Btu/24 hr 



)" 



5,700 Btu/hr 



PROBLEMS 

1. A cooler wall 10 ft by 18 ft is insulated with 
the equivalent of 3 in. of corkboard. Compute 
the heat gain through the wall in Btu/24 hr if the 
inside temperature is 37° F and the outside 
temperature is 78° F. Ans. 165,312 Btu/24 hr 

2. The north wall of a locker plant is 12 ft by 
60 ft and is constructed of 8 in. hollow clay tile 
insulated with 8 in. of corkboard. The locker 
plant is located in Houston, Texas and the 
inside temperature is maintained at 0° F. Deter- 
mine the heat gain through the wall in Btu/24 hr. 

Ans. 54,000 Btu/24 hr 

3. A cold storage warehouse in Orlando, 
Florida has a 30 ft by 50 ft flat roof constructed 
of 4 in. of concrete covered with tar and gravel 
and insulated with the equivalent of 4 in. of 
corkboard. If the roof is unshaded and the 
inside of the warehouse is maintained at 35° F, 
compute the heat gain through the roof in 
Btu/24 hr. Ans. 181,330 Btu/24 hr 

4. A small walk-in cooler has an interior volume 
of 400 cu ft and receives heavy usage. If the 
inside of the cooler is maintained at 35° F and 
the outside design conditions are 90° F and 60% 
relative humidity, determine the air change load 
in Btu/24 hr. Ans. 50,300 Btu/24 hr 

5. A frozen storage room has an interior volume 
of 2000 cu ft and is maintained at a temperature 



COOLING LOAD CALCULATIONS 163 



of -10° F. The usage is light and the outside 
design conditions (anteroom) are 50° F and 70 % 
relative humidity. Compute the air change load 
in Btu/24 hr. Arts. 16,100 Btu/24 hr 

6. Five thousand pounds of fresh, lean beef 
enter a chilling cooler at 100° F and are chilled 
to 38° F in 24 hr. Compute the chilling load in 
Btu/24 hr. Ans. 347,000 Btu/24 hr 

7. Five hundred pounds of prepared, packaged 
beef enter a freezer at a temperature of 36° F. 



The beef is to be frozen and its temperature 
reduced to 0° F in 5 hr. Compute the product 
load. Ans. 267,900 Btu/hr 

8. Fifty-five hundred crates of apples are in 
storage at 37° F. An additional 500 crates enter 
the storage cooler at a temperature of 85° F and 
are chilled to the storage temperature in 24 hr. 
The average weight of apples per crate is 60 lb. 
The crate weighs 10 lb and has a specific value 
of 0.6Btu/lb/°F. Determine the total product 
load in Btu/24 hr. Ans. l,679,80aBtu/24 hr 



II 

Evaporators 



1 1 — 1 . Types of Evaporators. As stated pre- 
viously, any heat transfer surface in which a 
refrigerant is vaporized for the purpose of 
removing heat from the refrigerated space or 
material is called an evaporator. Because of 
the many different requirements of the various 
applications, evaporators are manufactured in 
a wide variety of types, shapes, sizes, and 
designs, and they may be classified in a number 
of different ways, such as type of construction, 
operating condition, method of air (or liquid) 
circulation, type of refrigerant control, and 
application. 

1 1-2. Flooded and Dry-Expansion Evapora- 
tors. Evaporators fall into two general cate- 
gories, flooded and dry expansion, according to 
their operating condition. The flooded type is 
always completely filled with liquid refrigerant, 
the liquid level being maintained with a float 



valve or some other liquid level control (Fig. 
11-1). The vapor accumulating from the boiling 
action of the refrigerant is drawn off the top 
by the action of the compressor. The principal 
advantage of the flooded evaporator is that the 
inside surface of the evaporator is always 
completely wetted with liquid, a condition that 
produces a very high rate of heat transfer. The 
principal disadvantage of the flooded evapo- 
rator is that it is usually bulky and requires a 
relatively large refrigerant charge. 

Liquid refrigerant is fed into the dry-expan- 
sion evaporator by an expansion device which 
meters the liquid into the evaporator at a rate 
such that all the liquid is vaporized by the time 
it reaches the end of the evaporator coil (Fig. 
11-2). For either type, the rate at which the 
liquid is fed into the evaporator depends upon 
the rate of vaporization and increases or 
decreases as the heat load on the evaporator 
increases or decreases. However, whereas the 
flooded type is always completely filled with 
liquid, the amount of liquid present in the dry- 
expansion evaporator will vary with the load on 
the evaporator. When the load on the evapo- 
rator is light, the amount of liquid in the evapo- 
rator is small. As the load on the evaporator 
increases, the amount of liquid in the evaporator 
increases to accommodate the greater load. 
Thus, for the dry-expansion evaporator, the 
amount of liquid-wetted surface and, therefore, 
the evaporator efficiency, is greatest when the 
load is greatest. 

11-3. Types of Construction. The three 
principal types of evaporator construction are: 
(1) bare-tube, (2) plate-surface, and (3) finned. 




Liquid from 
receiver 

Float control 



Fig. Il-I. Flooded evaporator. 
Notice accumulator and float 
control. Circulation of the 
refrigerant through the coil 
is by gravity. The vapor accu- 
mulated from the boiling action 
in the coil -escapes to the top 
of the accumulator and is 
drawn off by the suction of the 
compressor. 



164 



EVAPORATORS l*S 



Bore-lube and plate-surface evaporators are 
sometimes dawned together as prime-surface 
evaporaion in thai the entire surface of both 
these types u more or leu in contact with the 
vaporiring refrigerant inside. With the finned 
evaporator, the refrigcran(<cafTying tube are 
the only prime surface. The fins themselves are 
not rilled wilh Tcfrigerani and are, therefore, 
only secondary hat transfer surfaces whose 
function J* (o pick up heat from the surrounding 
air and conduct it to the refrigerant-carrying 
tubes. 

Atthough prune-surface evaporate** of both 
the bare-tube and plait-surface types give salt* 
factory service on a wide variety of applications 
operating in any temperature range, they are 
most frequently applied to applications where 
the space temperature is maintained below 
34* F and frost accumulation on the evaporator 
surface cannot be readily prevented. Frost 
accumulation on prime-surface evaporaion 
does not affect the evaporator capacity to use 
extent thai it doc* on finned coils. Further' 
more, most prime surface evaporators, particu- 
larly the fiate-turface type, are easily cleaned 
and can be readily defrosted manually by either 
brushing or scraping off the frost accumulation. 
This can be accomplished without interrupting 
(be refrigerating process and jeopardizing the 
quality of the refrigerated product. 
(ML ftar*~Tub* Evaporator*. Bare- lube 
evaporators a re usually constructed of cither steel 
pipe or copper tubing. Steel pipe is used for 
luge evaporators and for evaporators to be 
employed with ammonia, whereas copper tubing 
it utilized in the manufacture of smaller evepo* 



Liquid *nm Rtfnetrtnt 

ntctiw j^Ro* contra* 




3 



3 



3> 



Mr Mb 

Fig. 11-5. Dry-wpvtslun aviporuer. Liquid r#- 
frlftrtftl refwrim prof r«tii inly u It flown through 
OOll Ifld l»*yet toil u i vapor. F«cler bulb cpnireU 
rata of now through tha orrflca at tha flow control. 




< U 'c 



(a) 




c 



T 



d6a 



M 



Flf. 11-3. Common dmgm far b*rt-tub* «H». 
(o> fin cignf eo+l. {&} C»*»l trwnewi* coil. 



niton intended for use with refrigerants other 
than ammonia. Bare-tube coils are available 
tn a number of sues, shapes, and designs, and 
are usually custom made lo the individual 
application. Common shapes for bare-tube 
coils are fiat zigzag and oval trombone, as 
shown in Fig. I!-), Spiral bare-tube coils are 
often employed for liquid chilling. 
11-5. Plate-Surface Evaporators. Plale-sur* 
face evaporators arc of several lypes. Some are 
constructed of two fiat sheets of metal so 
embossed and welded together as to provide a 
path for refrigerant flow between the two sheets 
(Fig. 11-4). This particular type of plate- 
■urfacc evaporator is widely uaed in household 
refrigerators and home freezers because it is 
easily cleaned, economical to manufacture, and 
can be readily formed into any one of the various 
shapes required (Fig, 11-5). 






(M PRINCIPLES OF REFRIGERATION 




Flf. MM. Standard lerpenllne plate evaporator, (Courtety Kold-Hold Division — Truuer Manufacturing, Inc.) 




■ ^^ r mmm 

Rg. 11-5. Soma typical iheocs «raJl*b!e in pfaie-typ* eviporiion. (Courtesy D«s- Product*. Inc.) 




(A) Outside jacket Of plate. Heavy, olnctrkally (E) Fitting whir* vacuum ll drawn and than permajv 
wddfid steel. Smooth surface, ontlv (tiled. 

(B) Continuous steel tubing through which refrlg- (F) Vacuum ip«« In dry pit to Spice in hold-over 
eratit passei. p.| K( contains lutectk solution under vacuum. No 
{C) Inlet from compressor, maintenance required due. CO i!urdy, limp In con- 
(D> Outfit to compressor. Copper connections for struct Jon. No moving parti; no th ing to wear or get 
all refrigerant* except ammonia where stiel cormee- out of order; no service neeusiry. 

lions ere vied. 

Fig. II-*, Plate-type evapc-retor, (Courtesy Dolt Refrigerating Company.} 



EVAPORATORS 



1*7 



Another type of plalc-surfacc evaporator 
consuls of formed tubing installed between two 
metal plates which arc welded together at the 
edges (Fig. It -6). In order to provide good 
thermal contact between the welded plated tnd 
the tubing carrying the refrigerant, live space 
between the plates is either filled with a eutcciic 
solution or evacuated so that the pressure of the 
atmosphere exerted on the outside surface of 
the plates holds the plates firmly against the 
tubing inside. Those containing the eutectic 
Solution are especially useful where a holdover 
capacity is required. Many are used on refriger- 
ated trucks, let such applications, the plates are 
mounted either vertically or horizontally from 
the ceiling or walls of the truck { Fig. 1 1-7) and 
are usually connected to a central plant refriger- 
ation system while the trucks are parked at the 
terminal during the night. The refrigerating 
capacity thui stored in the eutectic solution is 
sufficient to refrigerate the product during the 
oral day's operations. The temperature of the 



plates is controlled by the melting point of the 
eutectic solution. 

Plate-type evaporators may be used singly or 
in banks. Figure 11-8 illustrates how the plates 
can be grouped together for ceiling mounting in 
holding rooms, locker plants, freezers, etc. The 
plates may be manifolded for parallel flow of 
the refrigerant (Fig. 11-9) or they may be 
connected for series flow, 

Plate-surface evaporators provide excellent 
shelves in freezer rooms and similar applications 
(Fig. 11-10). They are also widely used a* 
partitions in freezers, frozen food display cases, 
Ice cream cabinets, uhIji fountains, etc. Plate 
evaporators are especially useful for liquid 
cooling installations where unusual peak load 
conditions are encountered periodically. By 
building up an ice bank on the surface of the 
plain during periods of light loads, a holdover 
refrigerating capacity is established which will 
help i be refrigerating equipment carry the load 
through the heavy or peak conditions Fig, 11-11), 




Pig, 1 1*1. Frtmitr plitu Iniullad In whottHlt Fc* crttm truck body, (Court aty Do4« H«tri| •r»tln| Company.) 



IfiS PRINCIPLES OF REFRIGERATION 




Flj. 11-$. P1»te b»nks employed in I o w MfnpcnLu r* itor-ifi roomt. (Ccmttmy Dqfa ftcfrt|enuifl| Company.) 



Since this allows Che use of smaller capacity 
equipment than would ordinarily be required 
by the peak load, a savings « affected in Initial 
cost and, usually, also in operating expenses, 
1 1-6. Finned Evaporators. Finned eoib in 
bare- tube coils upon which metal plates or fins 
have been insialled (Fig, 13-15), The fins, 
serving as secondary heat-absorbing surfaces, 
have the effect of increasing the over-all surface 
area of the evaporator, thereby improving its 
efficiency. With bare-lube evaporators, much 
of the air that circa kites over the coil passes 
through the open spaces between Ihe tubes and 
docs not come in contact with the cm I surface. 
When fins arc added to a coil, the fins extend 
out imo the open spaces be (ween the tubes and 
act as heat collectors. They remove heat from 
that portion of the air which would not ordin- 
arily conic in contact with the prime surface 
and conduct it back to the tubing, 

It is evident that to be effective the fins must 



be connected to the tubing in such a manner 
that good thermal contact between the fins and 
the tubing is assured. In some instances, the 
flni are soldered directly to [he i ubing. In others, 
the fins are slipped over ihe tubing and the 
tubing is expanded by pressure or some such 
means so that the fins bite imo the tube surface 
and establish good thermal contact. A variation 
of the latter method is to flare the fin hole 
slightly to allow the fin to slip over the tube. 
After the fin is installed, the flare is straightened 
and the fin is securely locked to the tube. 

Fin Ni« and spacing dej^eiid in part on the 
particular type of application for which the coil 
is designed, The size of the tube determines the 
size of the fin. Small tubes require small fins. 
As the size of the tube increases, the size of the 
fin may be effectively increased, Fin spacing 
varies from one to fourteen fins per inch, 
depending primarily on the operating tempera- 
ture of the coil. 



fVAfORATOAS 



16? 




Fl|» 1 1-*. Ptau bank with pluu minifeJdtd for ptraJM rrfriftnr* ftow. Plat** may ttto b« Kmn«t*d for 
ien« tlow. (Cduntiy Kold-Hdd Dickon— Trinwr H»nuf»ctur1ni, '"t-J 



Fig. tl-10. P1*» fnpQjiwn 
implayid u frtenr indvti. 
Not* thu plitM »r* imngtdi 
for Mhti ndrtftnni flow. 
{Courtesy KoJd-Hold Division 
— Trwwr Minuhcturinf , Jnc) 




170 PRINCIPLES OF AEFAKJEAATTON 




Ft t . 11.11. let-Cat Rtfrttar*- 
tion hoMov«r- capacity It etnb- 
li*h*d by buii^.ftf up ■ tank 
of l« on plite «aipenun, 
(Court**)* Dol* Rafrlfarilinf 

Cwnptflf.) 



Frost accumulation on air-cooling colls 
operating at low temperatures is unavoidable, 
and since any frost accumulation on finned coils 
tends to restrict the air passages between the 
lint and to retard air circulation through the 
coil, evaporators designed for tow temperature 
application! must have wide fin spacing (two 
or Ihree fins per inch) in order to minimize the 
danger of blocking air circulation. On the 
other hand, coils designed for air conditioning 
and other installation* where the coil Optra ICi 
it temperatures high enough so that no frost 
accumulates On the coil surface may have as 
many as fourteen fins per inch. 

When air circulation over finned coils is by 
gravity, it is important that the coil offer as little 
resistance lo air How as is possible; therefore, 
in general, fin spacing should be wider for 
natural convection coil* than for coils employing 
fans. 

It has been determined that a definite relation- 
ship exists between the inside and outside 
surfaces of an evaporator . Since external finning 
affects only the outside surface, I he addition of 
(ins beyond a certain limit will not materially 
increase the capacity of the evaporator. In 
fad. in some instances, excessive finning may 
actually reduce the evaporator capacity by 



restricting the air circuit son over the coil 
unnecessarily. 

Since their capacity is affected more by frost 
accumulation than any other type of evaporator, 
finned coils ore best suited to air-cooling applica- 
tion where the temperature is maintained above 
34' F. When tinned soils are used for low 
temperature operation, some means of defrost* 
ing the coil at regular inter val* must be provided. 
This may be accomplished automatically by 
several means which are discussed in another 
chapter , 

Because of the fins, finned coils have more 
surface area per unit of length and width than 
prime-surface evaporators and can therefore 
be built more compactly. Generally, a finned 
coil will occupy less space than either a bare- 
lube or plate-surface evaporator Of* the same 
capacity. Hut provides for a considerable 
savings in space and makes tinned coils ideally 
suited for use with funs a* In reed convection 
units. 

1 1-7. Evaporator Capacity. The capacity of 
any evaporator or cooling coil is the rate at 
which heat will pass th rough the evaporator 
walls from the refrigeruk-n ipiioe or product to 
the vaporizing liquid inside and is usually 
expressed in Blu per hour An evaporator 



selected for any specific application must always 
have sufficient capacity to allow the vaporizing 
refrigerant to absorb heat at the rate necessary 
to produce the required cooling when operating 
at the design conditions. 

Heat reaches the evaporator by all three 
methods of heat transfer. In air-cooling applica- 
tions most of the heat is carried to the evapo- 
rator by convection currents set up in the 
refrigerated space either by action of a fan or by 
gravity circulation resulting from the difference 
in temperature between the evaporator and the 
space. Too, some heat is radiated directly to 
the evaporator from the product and from the 
wall of the space. Where the product is in 
thermal contact with the outer surface of the 
evaporator, heat is transferred from the product 
to the evaporator by direct conduction. This 
is always true for liquid cooling applications 
where the liquid being cooled is always in 
contact with the evaporator surface. However, 
circulation of the cooled fluid either by gravity 
or by action of a pump is still necessary for 
good heat transfer. 

Regardless of how the heat reaches the out- 
side surface of the evaporator, it must pass 
through the walls of the evaporator to the 
refrigerant inside by conduction. Therefore, 
the capacity of the evaporator, that is, the rate 
at which heat passes through the walls, is deter- 
mined by the same factors that govern the rate 
of heat flow by conduction through any heat 
transfer surface and is expressed by the formula 

Q =A x U x D (11-1) 

where Q = the quantity of heat transferred in 
Btu/hr 
A = the outside surface area of the 
evaporator (both prime and finned) 
U = the over-all conductance factor in 
Btu/hr/sq ft of outside surface/" F D 
D = the logarithmic mean temperature 
difference in degrees Fahrenheit 
between the temperature outside the 
evaporator and the temperature of 
the refrigerant inside the evaporator 
11-8. U or Over-All Conductance Factor. 
The resistance to heat flow offered by the evapo- 
rator walls is the sum of three factors whose 
relationship is expressed by the following: 



1 _R L 1 

U~fi + K + f 



(11-2) 



EVAPORATORS 171 

where V = the over-all conductance factor in 

Btu/hr/sq ft/ F D 
ft = the conductance factor of the inside 

surface film in Btu/hr/sq ft of inside 

surface/" F D 
LjK = resistance to heat flow offered by 

metal of tubes and fins 
/„ =.the conductance factor of the out- 
side surface film in Btu/hr/sq ft of 

outside surface/" F D 
R = ratio of outside surface to inside 

surface 

Since a high rate of heat transfer through the 
evaporator walls is desirable, the U or conduct- 
ance factor should be as high as possible. 
Metals, because of their high conductance 
factor, are always used in evaporator construc- 
tion. However, a metal which will not react 
with the refrigerant must be selected. Iron, steel, 
brass, copper, and aluminum are the metals 
most commonly used. Iron and steel are not 
affected by any of the common refrigerants, 
but are apt to rust if any moisture is present in 
the system. Brass and copper can be used with 
any refrigerant except ammonia, which dis- 
solves copper. Aluminum may be used with 
any refrigerant except methyl chloride. Mag- 
nesium and magnesium alloys cannot be used 
with the fluorinated hydrocarbons or with 
methyl chloride. 

Of the three factors involved in Equation 
11-2, the metal of the evaporator walls is the 
least significant. The amount of resistance to 
heat flow offered by the metal is so small, 
especially where copper and aluminum are 
concerned, that it is usually of no consequence. 
Thus, the U factor of the evaporator is deter- 
mined primarily by the coefficients of con- 
ductance of the inside and outside surface 
films. 

In general, because of the effect they have on 
the inside and outside film coefficients, the value 
of U for an evaporator depends on the type 
of coil construction and the material used, the 
amount of interior wetted surface, the velocity 
of the refrigerant inside the coil, the amount 
of oil present in the evaporator, the material 
being cooled, the condition of the external 
surface, the fluid (either gaseous or liquid) 
velocity over the coil, and the ratio of inside 
to outside surface. 



172 PRINCIPLES OF REFRIGERATION 



Heat transfer by conduction is greater through 
liquids than through gases and the rate at which 
the refrigerant absorbs heat from the evaporator 
walls increases as the amount of interior wetted 
surface increases. In this respect, flooded 
evaporators, since they are always completely 
filled with liquid, are more efficient than the 
dry-expansion type. This principle also applies 
to the external evaporator surface. When the 
outside surface of the evaporator is in direct 
contact with some liquid or solid medium, the 
heat transfer by conduction to the outside 
surface of the evaporator is greater than when 
air is the medium in contact with the evaporator 
surface. 

Any fouling of either the external or internal 
surfaces of the evaporator tends to act as ther- 
mal insulation and decreases the conductance 
factor of the evaporator walls and reduces the 
rate of heat transfer. Fouling of the external 
surface of air-cooling evaporators is usually 
caused by an accumulation of dust and lint 
from the air which adheres to the wet coil 
surfaces or by frost accumulation on the coil 
surface. In liquid-cooling applications, fouling 
of the external tube surface usually results from 
scale formation and corrosion. Fouling of the 
internal surface of the evaporator tubes is 
usually caused by excessive amounts of oil in 
the evaporator and/or low refrigerant velocities. 
At low velocities, vapor bubbles, formed by the 
boiling action of the refrigerant, tend to cling 
to the tube walls, thereby decreasing the amount 
of interior wetted surface. Increasing the re- 
frigerant velocity produces a scrubbing action on 
the walls of the tube which carries away the oil 
and bubbles and improves the rate of heat flow. 
Thus, for a given tube size, the inside film coeffi- 
cient increases as the refrigerant velocity in 
creases. The refrigerant velocity is limited, how- 
ever, by the maximum allowable pressure drop 
through the coil and, if increased beyond a certain 
point, will result in a decrease rather than an in- 
crease in coil capacity. This depends to some 
extent on the method of coil circuiting and is dis- 
cussed later. It can be shown also that the con- 
ductance of the outside surface film is improved 
by increasing the fluid velocity over the outside 
surface of the coil. But, here again, in many 
cases the maximum velocity is limited, this time 
by consideration other than the capacity of the 
evaporator itself. 



Any increase in the turbulence of flow either 
inside or outside the evaporator will materially 
increase the rate of heat transfer through the 
evaporator walls. In general, internal turbu- 
lence increases with the difference in temperature 
across the walls of the tube, closer spacing of 
the tubes, and the roughness of the internal tube 
surface. In some instances, heat transfer is 
improved by internal finning. 

Outside flow turbulence is influenced by fluid 
velocity over the coil, tube spacing, and the 
shape of the fins. 

11-9. The Advantage of Fins. The advantage 
of finning depends on the relative values of the 
coefficients of conductance of the inside and 
outside surface films and upon R, the ratio of 
the outside surface to the inside surface. In any 
instance where the rate of heat flow from the 
inside surface of the evaporator to the liquid 
refrigerant is such that it exceeds the rate at 
which heat passes to the outside surface from 
the cooled medium, the over-all capacity of the 
evaporator is limited by the capacity of the 
outside surface. In such cases, the over-all 
value of U can be increased by using fins to 
increase the outside surface area to a point such 
that the amount of heat absorbed by the outside 
surface is approximately equal to that which 
can pass from the inside surface to the liquid 
refrigerant. 

Because heat transfer is greater to liquids than 
to vapors, this situation often exists in air- 
cooling applications where the rate of heat flow 
from the inside surface to the liquid refrigerant 
is much higher than that from the air to the 
outside surface. For this reason, the use of 
finned evaporators for air-cooling applications 
is becoming more and more prevalent. On the 
other hand, in liquid-cooling applications, since 
liquid is in contact with both sides of the evapo- 
rator and the rate of flow is approximately equal 
for both surfaces, barenube evaporators per- 
form at high efficiency and finning is usually 
unnecessary. In some applications, where fluid 
velocity over the outside of the evaporator is 
exceptionally high, the flow of heat to the outer 
surface may be greater than the flow from the 
inner surface to the refrigerant. When this 
occurs, the use of inner fins will improve evapo- 
rator capacity in that the amount of interior 
wetted surface is increased. Several methods 
of inner finning are shown in Fig. 11-12. 



EVAPORATORS 



173 





(a) 



(b) 



Fig. 11-12. Some methods of 
inner finning. 



11-10. Logarithmic Mean Temperature 
Difference. As illustrated in Fig. 11-13, the 
temperature of air (or any other fluid) decreases 
progressively as it passes through the cooling 
coil. The drop in temperature takes place along 
a curved line (A) approximately as indicated. 
Assuming that the temperature of the refrigerant 
remains constant, it is evident that the difference 
in temperature between the refrigerant and the 
air will be greater at the point where the air 
enters the coil than at the point where it leaves, 
and that the average or mean difference in 
temperature will fall along the curve (a) at a 
point somewhere between the two extremes. 
Although the value obtained deviates slightly 
from the actual logarithmic mean, an approxi- 
mate mean temperature difference may be 
calculated by the following equation: 

D JU-t T )^H-Q (U3) 

where D = the arithmetic mean temperature 
/„ = the temperature of the air entering 

the coil 
fj = the temperature of the air leaving 

the coil 
t r = the temperature of the refrigerant 

in the tubes 




(O 

For the values given in Fig. 11-13, the arith- 
metic mean temperature difference is 



D 



(40 - 20) + (30 - 20) 



= 15° F 



It must be remembered that the MTD as 
calculated by Equation 11-3 is slightly in error 
because of the curvature of the curved line A 
and would be the actual MTD only if the drop 
in air temperature occurred along a straight line, 
as indicated by the dotted line B in Fig. 11-13. 

The actual logarithmic mean temperature, 
which is the midpoint of the curved line A, is 
given by equation 

(f. - t r ) - (/ x - t r ) 



D = 



In 



it. - t r ) 

0l - t r ) 



(11-4) 



For the values given in Fig. 1 1 - 1 3, the logarith- 
mic mean temperature difference will be 



D = 



(40 - 20) - (30 - 20) 



In 



40-20 
30 -20 



JO 
ln2 



= — = 14.43° F 



The preceding calculations were made on the 
assumption that the refrigerant temperature 
remains constant. When this is not die case, 



174 PRINCIPLES OF REFRIGERATION 




Leaving air 
temperature-30 F 



Fig. 11-13. Mean temperature of air passing through 
evaporator. 



The velocity of the air passing over the coil 
has a considerable influence on both the value 
of U and the METD and is important in deter- 
mining evaporator capacity. When air velocity 
is low, the air passing over the coil stays in 
contact with the coil surface longer and is 
cooled through a greater range. Thus, the 
METD and the rate of heat transfer is low. As 
air velocity increases, a greater quantity of air 
is brought in contact with the coil per unit of 
time, the METD increases, and the rate of heat 
transfer improves. In addition, high air 
velocities tend to break up the thin film of 
stagnant air which is adjacent to all surfaces. 
Since this film of air acts as a heat barrier and 
insulates the surface, its disturbance increases 
the conductance of the outside surface film and 
the over-all value of U improves. 



t r will have two values. This condition is 
discussed in another chapter. 

The log mean temperature difference here- 
after called mean effective temperature difference 
(METD), may also be determined from Table 
11-1. 

11-11. The Effect of Air Quantity on 
Evaporator Capacity. Although not a part of 
the basic heat transfer equation, there are other 
factors external to the coil itself which greatly 
affect coil performance. Principal among these 
are the circulation, velocity, and distribution of 
air in the refrigerated space and over the coil. 
These factors are closely related and in many 
cases are dependent one on the other. 

Except in liquid cooling and in applications 
where the product is in direct contact with the 
evaporator, most of the heat from the product 
is carried to the evaporator by air circulation: 
If air circulation is inadequate, heat is not 
carried from the product to the evaporator at a 
rate sufficient to allow the evaporator to per- 
form at peak efficiency. It is important also 
that the circulation of air is evenly distributed 
in all parts of the refrigerated space and over 
the coil. Poor distribution of the circulating 
air results in uneven temperatures and "dead 
spots" in the refrigerated space, whereas the 
uneven distribution of air over the coil surface 
causes some parts of the surface to function less 
efficiently than others and lowers evaporator 
capacity. 




CoflA 




CoilB 



A 



■Air 



W 



CoilC 

Fig. 1 1-14. Coils B and C both have twice the surface 
area of coil A. Coil C has twice the face area of coil A 
or coil B. 



EVAPORATORS 175 



11-12. Surface Area. Equation 11-1 indicates 
that the capacity of an evaporator varies 
directly with the outside surface area. This 
is true only if the U factor of the evaporator 
and the METD remains the same. In many 
cases, the value of U and the METD are affected 
when the surface area of the evaporator is 
changed. In such cases, the capacity of the 
evaporator does not increase or decrease in 
direct proportion to the change in surface area. 
To illustrate, in Fig. 11-14, coils B and C each 
have twice the surface area of coil A, yet the 



ing the number of rows as in coil B, the METD 
will be decreased and the increase in capacity 
will not be nearly as great as when the surface 
area is increased as in coil C. For the same 
total surface area, a long, wide, flat coil will, in 
general, perform more efficiently than a short, 
narrow coil having more rows depth. However, 
in many instances, the physical space available 
is limited and compact coils arrangements must 
be used. In applications where it it is permis- 
sible, the loss of capacity resulting from increas- 
ing the number of rows can be compensated for 



3rd 
Row 



Fig. 11-15. Air temperature 
drop across typical three-row 
cooling coil. 



Leaving -<- 

Air -*r- 



Temperature -<- 



30*-«- 



increase in capacity over the capacity of coil A 
will be greater for coil C than for coil B. Pro- 
vided the air velocity is the same (the total 
quantity of air circulated over coil C must be 
twice that circulated over coil A), the METD 
across C will be exactly the same as that across 
A and the capacity of C will therefore be twice 
the capacity of coil A. 

Figure 11-15 shows how the METD is af- 
fected when the surface area of the coil is 
increased by increasing the number of rows 
(depth). Note that the drop in air temperature 
is much greater across the first row and dimin- 
ishes as the air passes across each succeeding 
row. This is accounted for by the fact that the 
temperature difference between the air and the 
refrigerant is much greater across the first row, 
becomes less and less as the temperature of the 
air is reduced in passing across each row, and 
is least across the last row. It is evident then 
that the rate of heat transfer is greater for the 
first row and that the first row performs the 
most efficiently. For this reason, if the surface 
area of coil A in Fig. 11-14 is doubled by increas- 





2nd 




Row 







Air 














to 










32° 


M 




0) 


^ 


CC 










Air 


C\J 









1st 






Row 




■< 

Air 


C 

2 


-< 












35° 


no 




^ 








o 




Air 


CM 











Entering 
Air 

Temperature 
40* 



Coil 



to some extent by increasing the air velocity over 
the coil. Too, in some applications, the use of 
deep coils is desirable for the purpose of 
dehumidification. 

11-13. Evaporator Circuiting. It was demon- 
strated in Chapter 8 that excessive pressure drop 
in the evaporator results in the suction vapor 
arriving at the suction inlet of the compressor 
at a lower pressure than is actually necessary, 
thereby causing a loss of compressor capacity 
and efficiency. 

To avoid unnecessary losses in compressor 
capacity and efficiency, it is desirable to design 
the evaporator so that the refrigerant experi- 
ences a minimum drop in pressure. On the 
other hand, a certain amount of pressure drop 
is required to flow the refrigerant through the 
evaporator, and since velocity is a function of 
pressure drop, the drop in pressure must be 
sufficient to assure refrigerant velocities high 
enough to sweep the tube surfaces free of vapor 
bubbles and oil and to carry the oil back to the 
compressor. Hence, good design requires that 
the method of evaporator circuiting be such 



176 PRINCIPLES OF REFRIGERATION 




Refrigerant 
out 



Fig. 11-16. Evaporator with one series refrigerant 
circuit. 



that the drop in pressure through the evaporator 
is the minimum necessary to produce refrigerant 
velocities sufficient to provide a high rate of heat 
transfer and good oil return. 

In general, the drop in pressure through any 
one evaporator circuit will depend upon the size 
of the tube, the length of the circuit, and the 
circuit load. By circuit load is meant the time- 
rate of heat flow through the tube walls of the 
circuit. The circuit load determines the quan- 
tity of refrigerant which must pass through the 
circuit per unit of time. The greater the circuit 
load, tiie greater must be the quantity of 
refrigerant flowing through the circuit and the 
greater will be the drop in pressure. Hence, 
for any given tube size, the greater the load on 
the circuit, the shorter the circuit must be in 
order to avoid excessive pressure drop. 

Evaporators having only a single series 
refrigerant circuit, such as the one illustrated 
in Fig. 11-16, will perform satisfactorily within 
certain load limits. When the load limit is 
exceeded, the refrigerant velocity will be in- 
creased beyond the desired range and the 
pressure drop will be excessive. 

Notice that the refrigerant enters at the top 
of the evaporator as a liquid and leaves at the 
bottom as a vapor. Since the volume of the 
refrigerant increases as the refrigerant vaporizes, 
the refrigerant velocity and the pressure drop 
per foot increase progressively as the refrigerant 
travels through die circuit, and are greatest at 
the end of the coil where the refrigerant is 100 % 
vapor. 

The excessive pressure drop occurring in the 
latter part of a single series circuit evaporator 



can be eliminated to a great extent by splitting 
the single circuit into two circuits in the lower 
portion of the evaporator (Fig. 11-17). When 
this is done, the refrigerant travels a single series 
path until the refrigerant velocity builds to the 
allowable maximum, at which time the circuit 
is split into two parallel paths for the balance of 
the travel through the evaporator. This has the 
effect of reducing the refrigerant velocity in the 
two paths to one-half the value it would have 
without the split, and the pressure drop per foot 
is reduced to one-eighth of the value it would 
have in the lower part of the evaporator with a 
single single series circuit.* This, of course, 
will permit greater loading of the coil without 
exceeding the allowable pressure drop. At the 
same time, the velocity in all parts of the coil 
is maintained within the desirable limits so 
that the rate of heat transfer is not unduly 
affected. 

Another method of reducing the pressure drop 
through the evaporator is to install refrigerant 
headers at the top and bottom of the evaporator 
so that the refrigerant is fed simultaneously 
through a multiple of parallel circuits (Fig. 
11-18). However, this arrangement is not too 
satisfactory and is not widely used. While the 
pressure drop through the evaporator is low, 
this method of circuiting ordinarily results in 




Fig. 11-17. Evaporator with split refrigerant circuit. 

* Pressure drop increases as the square of the 
velocity. Reducing the velocity to one-half reduces 
the pressure drop to one-quarter of its original value. 
Then, since the length of each parallel branch is 
only one-half the length of a single circuit, the drop 
in pressure in the lower portion of the split coil is 
only one-eighth of the single circuit value. 



EVAPORATORS 177 



reducing the refrigerant velocity below the 
desired minimum so that the inside film coefficient 
and the rate of heat transfer are also low. 
Another disadvantage of this type of circuiting 
is that the loading of the circuits is uneven. Since 
the temperature difference between the air 
passing over the coil and the refrigerant in the 
tubes is much greater across the first circuit (first 
row) than across the last circuit (last row), the 
loading of the first circuit is much greater than 
the loading of the last circuit. Hence, the 
refrigerant velocity and the drop in pressure 
through the several circuits are uneven and a 
large portion of the coil operates inefficiently. 
This criticism can be applied to some extent 
also for the circuit arrangements in Figs. H-16 
and 1 1-17. In all three arrangements, the lower 
portion of the evaporator will not perform as 
effectively as the upper portion because wetting 
of the internal tube surface will not be as great 
in the lower portion. This is because the refrig- 
erant in the lower portion contains a high 
percentage of vapor, whereas in the upper 
portion the refrigerant is nearly all liquid. 

It is for this reason also that the outside 
surface temperature of the coil is always lowest 
near the refrigerant inlet and highest near the 
outlet, in spite of the fact that the saturation 
temperature of the refrigerant is lowest at the 
outlet due to the drop in pressure through the 
coil. 

The circuit arrangement shown in Fig. 11-19 
is very effective and is widely used, particularly 
when circuit loading is heavy, as in the case of 



Refrigerant 

out V^ 



Refrigerant 

in 





Circuits 



Fig. 11-18. Four-circuit evaporator with refrigerant 
headers on both inlet and outlet. Crossflow of air 
and refrigerant results in uneven circuit loading. 



Refrigerant 

in 

Refrigerant 
distributor 



. > Sk. Refrigerant 
out 

Fig. 11-19. Evaporator with refrigerant distributor 
and suction header. Notice counterflow arrangement 
for refrigerant and air. 



an air conditioning coil where the temperature 
differential between the refrigerant and the air 
is large and where external finning is heavy. 
Notice that the air passes in counterflow to the 
refrigerant so that the warmest air is in contact 
with the warmest part of the coil surface. This 
provides the greatest mean temperature differ- 
ential and the highest rate of heat transfer. 
Notice also that loading of the circuits is even. 
The number and length of the circuits that such 
a coil should have are determined by the size 
of the tube and the load on the circuits. 

For the multipass, headered evaporator, the 
arrangement shown in Fig. 11-20 is much more 
desirable than that shown in Fig. 11-18. Coun- 
terflowing of the air and the refrigerant increases 
the METD and permits more even loading of 
the circuits. 

1 1-14. Use of Manufacturer's Rating Tables. 
The mathematical evaluation of all the factors 
which influence evaporator capacity is usually 
impractical and in many cases impossible. For 
the most part, evaporator capacities must be 
determined by actual testing of the evaporator. 
The results of such tests are contained in the 
rating tables published by the various evapo- 
rator manufacturers. 

The method of rating evaporators varies 
somewhat with the type of evaporator and with 
the particular manufacturer involved. How- 
ever, the various rating methods are very similar 
and most manufacturers include, along with the 
evaporator rating tables, instructions as to how 
to use the ratings. In most cases, where the 



178 PRINCIPLES OF REFRIGERATION 




Fig. 11-20. Counter-flow of 
refrigerant and air results in 
more even circuit loading and 
a higher mean temperature 
differential. Compare this 
arrangement with the cross- 
flow arrangement in Fig. 11-18. 



evaporators are rated in accordance with ASRE 
Standards, the capacity data are reliable and 
are for operating conditions as normally en- 
countered. 

The selection of evaporators from manufac- 
turer's rating tables is relatively simple once the 
conditions at which the evaporator is to operate 
are known. Typical evaporator rating tables, 
along with methods of evaporator selection, are 
discussed at the end of the chapter. 
11-15. Evaporator TD. One of the most 
important factors to consider in selecting the 
proper evaporator for any given application 
is the evaporator TD. Evaporator TD is 
defined as the difference in temperature between 
the temperature of the air entering the evapo- 
rator and the saturation temperature of the 
refrigerant corresponding to the pressure at the 
evaporator outlet.* 

Although more exact methods of rating 
evaporators are necessary in order to select 
evaporators for air conditioning applications 
and some product storage applications where 
space temperature and humidity are especially 
critical, ratings for most evaporators designed 
for product cooling applications are based on 
evaporator TD. 

The relationship between evaporator capacity 
and evaporator TD is shown by the curve in 
Fig. 11-21. Notice that the capacity of the 
evaporator (Btu/hr) varies directly with the 
evaporator TD. That is, if an evaporator has a 
certain capacity at a 1 ° F TD, it will have exactly 
ten times that capacity if the TD is increased to 

* ASRE Standard 25-56, Methods of Rating Air 
Coolers For Refrigeration. 



10° F, provided that all other conditions are the 
same.t 

It is evident that a coil with a relatively small 
surface area operating at a relatively large TD 
can have the same capacity as another coil 
having a larger surface area but operating at a 
smaller TD. Thus, insofar as Btu per hour 
capacity alone is concerned, a small coil will 
have that same refrigerating effect as a larger 
one, provided that the TD at which the small 
coil operates is greater in proportion. However, 
it will be shown in the following sections that 
the temperature difference between the evapo- 
rator and the refrigerated space has considerable 
influence on the condition of the stored product 
and upon the operating efficiency of the entire 
system, and is usually, therefore, the determining 
factor in coil selection. Before an evaporator 
can be selected, it is necessary to first determine 
the TD at which it is expected to function. Once 
the desired temperature difference is known, an 

t Care should be taken not to confuse METD 
with evaporator TD. According to Equation 11-1, 
the Btu per hour capacity of any given evaporator 
(whose U factor and surface area are fixed at the 
time of manufacture) varies directly with the 
METD. However, assuming that the refrigerant 
temperature and all else remains constant, the 
METD between the air passing over the evaporator 
and the refrigerant in the evaporator will vary 
directly with the temperature of the air entering the 
evaporator. That is, if the temperature of the air 
entering the evaporator increases, the METD 
increases. Hence, the METD varies in proportion 
to the evaporator TD and, therefore, the capacity 
of the evaporator also varies in proportion to the 
evaporator TD. 



evaporator having sufficient surface area to 
provide the required cooling capacity at the 
design TD can be selected. 
11-16. The Effect of Coil TD on Space 
Humidity. The preservation of food and other 
products in optimum condition by refrigeration 
depends not only upon the temperature of the 
refrigerated space but also upon space humidity. 
When the humidity in the space is too low, 
excessive dehydration occurs in such products 
as cut meats, vegetables, dairy products, flowers, 
fruits, etc. On the other hand, when the humid- 
ity in the refrigerated space is too high, the 
growth of mold, fungus, and bacteria is encour- 
aged and bad sliming conditions occur, particu- 
larly on meats and especially in the wintertime. 
Space humidity is of little importance, of course, 
when the refrigerated product is in bottles, cans, 
or other vapor-proof containers. 

The most important factor governing the 
humidity in the refrigerated space is the evapo- 
ratorTD.* The smaller the difference in tempera- 
ture between the evaporator and the space, the 
higher is the relative humidity in the space. 
Likewise, the greater the evaporator TD, the 
lower is the relative humidity in the space. 

When the product to be refrigerated is one 
that will be affected by the space humidity, an 
evaporator TD that will provide the optimum 
humidity conditions for the product should be 
selected. In such cases, the evaporator TD is 
the most important factor determining the 
evaporator selection. The design evaporator 
TD required for various space humidities is 
given in Table 11-2 for both natural-convection 
and forced-convection evaporators. 

In applications where the space humidity is 
of no importance, the factors governing evapo- 
rator selection are: (1) system efficiency and 
economy of operation, (2) the physical space 
available for evaporator installation, and (3) 
initial cost. 

11-17. The Effect of Air Circulation on 
Product Condition. As stated previously, 
circulation of air in the refrigerated space is 
essential to carry the heat from the product to 
the evaporator. When air circulation is inade- 

* Some of the other factors which influence the 
space humidity are: air motion, system running 
time, type of system control, amount of exposed 
product surface, infiltration, outside air conditions, 
etc. 



EVAPORATORS 17? 

quate, the capacity of the evaporator is de- 
creased, the product is not cooled at a sufficient 
rate, the growth of mold and bacteria is encour- 
aged, and sliming occurs on some products. On 
the other hand, too much air circulation can be 
as detrimental as too little. When the circula- 
tion of air, is too great, the rate of moisture 
evaporation from the product surface increases 
and excessive dehydration of the product 
results. Excessive dehydration can be very 
costly in that it causes deterioration in product 
appearance and quality and shortens the life 
of the product. Furthermore, the loss of weight 
resulting from shrinkage and trimming is a 
considerable factor in dealer profits and in the 
price of perishable foods. 

The desired rate of air circulation varies with 
the different applications and depends upon the 
space humidity, the type of product, and the 
length of the storage period. 

With respect to product condition, air circula- 
tion and space humidity are closely associated. 
Poor air circulation has the same effect on the 
product as high humidity, whereas too much air 
circulation produces the same effect as low 
humidity. In many instances, it is difficult to 
determine whether product deterioration in 
a particular application is caused by faulty air 
circulation or poor humidity conditions. For 
the most part, product condition depends upon 
the combined effects of humidity and air 
circulation, rather than upon the effect of either 
one alone, and either of these two factors can 
be varied somewhat, provided that the other is 
varied in an off-setting direction. For example, 
higher than normal air velocities can be used 
without damage to the product when the space 
humidity is also maintained at a higher level. 

The type of product and the amount of 



35 
£ 30 

a§2 5 

l' E 15 

uj 5 


























































































































































































5* 



10* 



15* 20* 25* 30* 35* 
Evaporator TD 



Fig. 11-21. Variation in evaporator capacity with 
evaporator TD. 



180 PRINCIPLES OF REFRIGERATION 



exposed surface should be given consideration 
when determining the desired rate of air circula- 
tion. Some products, such as flowers and 
vegetables, are more easily damaged by excessive 
air circulation than others and must be given 
special consideration. Cut meats, since they 
have more exposed surface, are more susceptible 
to loss of weight and deterioration than are 
beef quarters or sides, and air velocities should 
be lower. On the other hand, where the product 
is in vapor-proof containers, it will not be af- 
fected by high velocities and the rate of air 
circulation should be maintained at a high level 
to obtain the maximum cooling effect. 

Recommended air velocities for product 
storage are given in Tables 10-10 through 10-13. 
11-18. Natural Convection Evaporators. 
Natural convection evaporators are frequently 
used in applications where low air velocities 
and minimum dehydration of the product are 
desired. Typical installations are household 
refrigerators, display cases, walk-in coolers, 
reach-in refrigerators, and large storage rooms. 

The circulation of air over the cooling coil by 
natural convection is a function of the tempera- 
ture differential between the evaporator and the 
space. The greater the difference in temperature, 
the higher the rate of air circulation. 

The circulation of air by natural convection 
is greatly influenced by the shape, size, and 
location of the evaporator, the use of baffles, 
and the placement of the stored product in the 
refrigerated space. Generally, shallow coils 
(one or two rows deep) extending the entire 



length of the cooler and covering the greater 
portion of the ceiling area are best. As the 
depth of the coil is increased, the coil offers 
greater resistance to the free circulation of air 
and the METD is thereby decreased with a 
resulting decrease in the coil capacity. Since 
cold air is denser than warm air and tends to 
fall to the floor, evaporators should be located 
as high above the floor as possible, but care 
should be taken to leave sufficient room between 
the evaporator and the ceiling to permit the free 
circulation of air over the top of the coil. 

For coolers less than 8 ft wide, single, ceiling- 
mounted evaporators are frequently used. When 
the width of the cooler exceeds 8 ft, two or more 
evaporators should be used. In coolers where 
there is not sufficient head room to permit the 
use of overhead coils, side-wall evaporators 
may be used. If properly installed, these will 
function with approximately the same efficiency 
as overhead coils. Typical overhead and side- 
wall installations for large storage rooms are 
shown in Figs. 11-22 and 11-23, respectively. 

In small coolers, baffles are used with natural 
convection coils to assure good air circulation. 
The baffles are installed in such a manner that 
they aid and direct the free flow of air over the 
coil and throughout the refrigerated space. The 
cold and warm air flues should each have an 
area approximately equal to one-sixth or one- 
seventh of the floor area of the cooler. Assum- 
ing that the flues extend the full length of the 
fixture, the width of the flues will then be 
proportional to the width of the cooler. Since 




Fig. 11-22. Overhead installa- 
tion of natural convection 
evaporator. Evaporator has 
cast aluminum fins. (Courtesy 
Detroit Ice Machine Company.) 



EVAPORATORS 



III 



Ftf. II.23L Sid* will Installa- 
tion of naturil tonvtalon 
enpantor. (County On role 
Ic* Miehmt Company ) 




warm Air has a greater specific volume than 
eoJd air. some manufacturers recommend thai 
the warm air flue tx a little larger than the cold 
air flue. In Fig. 1 1-24, the width or the cold air 
Due it equal to W/7, whereas the width of the 
warm air flue is equal to W','6. The distance M) 
from the coil to the ceiling should be approxi- 
maldy equal to the width of the watm air flue 
and never less than 3 in. Vertical side baffles 
should extend approximately t in. above and 




Hf. 11-24. Tvpkal baflla amngamaflt for natunl 
conractlon end I. 



3 (o 4 in. beto* the coil. The horizontal baffles 
or coil decks should slope ) to 2 in. per foot to 
give direction to the cold ajr flow and lo drain 
on" the condensate- Also, I he coil decks must 
be insulated so that moisture does not condone 
on the undersurfcee of the deck and drip off on 
the product. The dimension (if) is 4 to 7 in., 
whereas (O is usually 1 to 4 in, 
IMf-Coi1-ajid-Bafr1eAii*mbll«s. The avail- 
ability of radory-buill coJI-and-bafrk assemblies 
has practically eliminated the need for the 
custom building of baffles on the job. A typkal 
ready built coil-and-bafflc assembly is shown 
in Fig. 11-25. Since these assemblies are 
available in a wide variety or sizes and combina- 
tion (an Table R-l), they can be readily 
applied to almost any natural convection 
application. 

11-20. Hating and Select I on uf Natural Con- 
vection Evaporators, Basic capacity ratings 
Tor natural convection evaporators, both prime 
surface and finned, arc normally given in Btu/hr/ 
* F TD. However, in some instances, where it 
will simplify evaporator selection, capacity 
ratings are given for TD's other than I • F. 

For the coil -and- baffle assemblies mentioned 
in the preceding section, the ratings given arc 
per inch of finned length. For bare pipe evapo- 
rators, the ratings given are per square foot of 
externa] pipe surface, although in some instances 
bare pipe evaporators are rated per lineal foot 
of pipe. 

Ratings for plate evaporators are given per 
squire foot of plate surface. Both sides of the 
plate arc considered when computing the area 



I K. MUNCltt.ES OF REfRlGEIUTlON 



Ft|. 11*35, Nitunl «cn»«. 

(Court -rty Dunhjm- faith. Int.) 




of the plate, Frequently, ratings Tor plate 
evaporators apply to an entire plate or to a 
specific group or combination or plats, 

Tvp^-.ii eating data f^? hhjcth typp of 
■mural convection evaporators are given in 
Tables R-l through R-7. The use or these 
rating data in the selection of the various types 
of evaporators is best illustrated through the 
use or a series of examples. 

Enamels J M. Select i natural convection 
cou-and-baflfe assembly (Fig. 1 J -25 and Table 
R-l) for the vegetable storage cool? in Example 
KM I. 

Solution. Since the capacity ratings for this 
type of evaporator are given in Btufhr/T TD/in. 
of finned length, the required evaporator capa- 
city must be reduced to this value before a 
selection can be made from the rating (able, 
Abo. recall that a natural convection evapora- 
tor should extend almost the full length of the 
cooler in order to assure adequate air circulation 
around the product. 

From Example 10-11. 
inside dimension of cooler 

Required evaporator 
capacity (average hourly 
cooling load) 

From Table 10-11. de- 
sired space humidity for 
mixed vegetable storage 

From Table 1 1-2, design 
evaporator TO required 
forS7%RH 



— 17 ft x9A 

— 8500Biu/hr 
*• Approx. 87% 



To determine approxi- 
mate finned length required : 
Over-all length of 
cooler (inside) — 17 ft 

Allowing I ft on each 
end of evaporator for 
working space, the 
approximate over-all 
length of the evapo- 
ratori4(«cFig.Jl-16) 
(I7fl-2ft) -|5ft 

According to the 
manufacturer's speci- 
fications {Table ft-1). 
the over-all length of 
the evaporator is 7 in 
longer than the actual 
finned length. Hence, 
the approximate 

finned length desired - 14 ft 3 in. 
is {15 ft -7 in.) or 171 in. 

To determine the require! capacity in Biu/hr/ 
*FTD,in. finned length: 
Required evaporator 
capacity per * F TD 

Total c* J|H>rutor capacity 

DewgiiL...!]«ralorTD 
8500 Blu hr 
14* Fin 
- e0 l 7B•u;hr.■ , FTD 
Required capacity 
{Biu/hr/*FTD/ineh 

m Rcquirett < o pacity per *FTD 

Dei i rcdTtnned length 
*07 Biu.hr/" F TD 



171 in. tinned length 
- 3.35 Btu/hrf F/ineh 



EVAPORATORS IS) 



Became or the width of the cooler, a two- 
tcciicti evaporator will give the best results. 
Reference to Tabic R-l indicates that Model 
#PK-I6 with two fin* per inch (J -in. fin spacing) 
has ■ capacity ofl 65 Biu/hr/ 1 F/in. 

Using this mode) evaporator, the required 
finned length is 



J,*5 BtWnr/' F/in. 



- lAfun 




The overall length of the evaporator is 173 
in. (166 in. + 7 in.) and. since the overall 
length of the cooler inside is 204 in., the 
dc*rancc between the evaporator and the cooler 
wall at each end is 15 5 in. 

The width of the evaporator should always 
be checked a gains I the width of the cooler lo be 
sure that the evaporator can be installed in the 
space in accordance with the manufacturer's 
recommendations. For evaporators of this 
type, the manufacturer recommends that the 
side of the evaporator be not less than 6 in. nor 
more than 12 in. from the cooler wall {installa- 
tion dimension A of Table R-l) and that the 
distance between the two sections of the evapo- 
rator (dimension C) be not less than 6 in. nor 
more than 8 ft. 

The maximum allowable evaporator width 
can be determined by subtracting the minimum 
of dimensions A and C from the inside width of 
the cooler, viz : 

Maximum width of evaporator •- Inside 
width of cooler - {A + C + Q. In this par- 
ticular case, the maximum allowable width of the 
evaporator is 

108 in. - (6 in. +6 in. + 6 in.) - 90 in. 

Since the combined width of the two sections 
of evaporator. Model #PK-16, is only 36 in. 



fit. W-tt , Arnnf«rn«ni of natural (on**ct»on 
twtporttori in itortj* COoUr (i*» Example 1 1. 1). 



(IS in. x 2), the evaporator is suitable for the 
cooler. A logical arrangement of ihc two 
evaporator sections in the cooler is shown in 
Fig 11-27. 

To order the evaporator, specify the modet 
number, fin spacing, and finned length, viz: 
Model #PK- 1 6.3-166 in 

Note. To avoid excessi vely long evaporators, 
which are inconvenient to ship and install, a 
multiple of evaporators should be used in large 
coolers. Typical arrangements Tor large coolers 
are shown in Fig. 11-28. These arrangements 
are also suitable For plate banks. 

ExampJ* 11*2, Using Table R-3, select 
plate evaporators (banks) for ceiling installation 
in the locker room of Example 10-16, To assure 
good air circulation in the locker room, select 
evaporators for a 15* FTD, 

Solution, Analysis of room dimensions 
(30.5 ft x 16 ft) indicates that four to six evapo- 
rators (two or three banks installed end-to- 
end over each aisle will be needed to provide 
good ceiling coverage and adequate air distri- 
bution. Reference to Table R-3 will show that 
plate banks are available in stock lengths of 
108 in. (9ft) and 144 in, (lift). Three banks 
108 in. long (a total of 27 ft) or two banks 144 
in, long (a total of 24 ft) could be installed 
end-to-end over each aisle and allow adequate 
working clearance at the ends. 

Since the banks are already rated at the design 
TD of 15* F, the ratings can be used directly and 
the required capacity per bank can be deter- 
mined by dividing the total hourly cooling load 
by the desired number of plate banks, viz: 

Hf> 1 1*1*, Arr»mj«m«ni of mmuni eonvteiion Required capacity _ Total cooling load (Btu/hr) 
tviponton In ttorij* coolir («■ Exampi* 1 t-|). per bank (Btu/hr) Number of banks desired 



inftdt hngei of coota?— I r ■ 




IB4 PRINCIPLES OF REFRIGERATION 



In this instance, the capacity required per bank 
is 



or 



24,650 Btu/hr 
6 buna 

24,650 Btu/hr 
4 banks 



-4lWBtWhr/13*FTD 



= 6l62Blu/hr/)5' FTD 



Referring to Table R-3, we see thai plate 
bank. Model #5-1 210B-B, has a capacity of 4320 
Btu/hr at a 15" F TD when operating below 
32° F (frosted). This will permit good coverage 
of the ceiling and at the same time allow suffi- 
cient working space ui the ends of the banks (see 
Fig- ll-29>. 

In ordering the evaporators, specify refrig- 
cram and the type of connections desired 
(series or manifold), viz : 

Mode) #6-J2l44-B, series connected for 
Refrigerant- 12. Notice {Table R-J) that the 
manufacturer specifies that two refrigerant flow 
controls should be used with each bank for 
Refrigeran 1-12, whereas only one a needed when 
ammonia is the refrigerant. 

Example 11-3, From Table R-4, select a 
plate-stand assembly for the freezing cabinet in 
Example 10-16. The inside dimensions of (be 
cabinet are 28 in. x 90 in. and the freezing load 
is 4590 Btu/hr. Base plate selection on 10 F 
TD. 



V 




Fig. 11-26. Typical arrange rrmnfi tor natural con- 
vection evaporator* In lirjt cooten. 



Solution. Inspection of Tabte R-4 will show 
that the ratings of the p I li : ._- stand assemblies 
an based on a J 5° F TD. Since the design TD 
in this instance is only 10 F. it is necessary to 
determine (he capacity the plate must have at a 
IV I TD in order to have the desired design 
TD of 10- F. This is accomplished by dividing 
the average hourly load by the design TD of 
10* F and then multiplying by the rating TD of 
15* F, vis; 



4590 Btu/hr x 13' F 
10" F 



6885 Btu/hr 



Thus, it is determined that an evaporator having 
a capacity of 6885 Btu/hr ai a 15 ' F TD will 
have I he desired capacity of 4590 Btu/hr at the 
design TD of ID* F. Th* value can then be 
used to select the plate stand directly from the 
rating table. 

By referring to Tabic R-4, plate stand, 
Model #1-72%*-%. which is approximately 26 in. 
wide and 88 in. long (including piping connec- 
tions), will (it the freeaorr cabinet and has a 
capacity of 7140 Btu/hr ai a 15" F TD (4760 
Btu/hr at 10* F TD>, This provides a small 
safety factor and is therefore satisfactory for 
the application. 

Alternate Solution. A rule of thumb used in 
selecting plate freezer? for locker plant appli- 
cations is to allow 0.5 Sq ft of plate surface for 
each locker. 

Allowing 0.5 sq ft of plate surface per locker 
per day, the plate surface required la this 
instance is 

353 lockers x 0.5 = 1 7fi.S sq ft 

By referring to Table R-2. the plate which best 
flu the cabinet. Model #22S4 (22 in. x 84 in.), 
has a surface area (both sides of plate) of 27.24 
sq ft per plate. Therefore, seven pistes of 
this size are required. Seven plates have a total 
capacity of 7140 Btu/hr ai a I5 5 F TD or 4760 
Btu/hr at a 10° F TD. 

Example 11-4. Assuming a space tem- 
perature of 0* F and a refrigerant temperature 
of - 1 T F (1 7" F evaporalor TD), determine the 
lineal feet of 1 i in. iron pipt required to handle 
the cooling load on the locker room in Example 
10-16. 

Solution, By referring to Table R*6, the 
capacity per square foot of pipe (outside surface) 
at the given conditions is 1.5 Btu/hr/ F TD, 
To determine the square feet of pipe surface 
required, divide the average hourly load by the 



EVAPORATORS IIS 



Flf. 1 1 -It, Inttillillon of pliie 
tanks in locker ptint (sm 
Eximpl* 11-3.) 







— 305" 




* 


|£ — 


*— ^r^ 


** r tl IT 


g* 


t\ 


" 








1 ^^^^^^^^^ ' 



*-r 



capacity per square foot per degree TD and by 
die design TD, viz: 

Opacity required (Btu/hr) 
Pipe capacity (Btu/hr/iq ft/' F) x TD 

— Pipe surface I sq ft) 

In this instance, the square feet of pipe surface 
required is 

24,650 Btu/hr 

1.3 x 17 ™>»M" 

By referring to Table R-7, 2.1 lineal ft of I J in. 
pipe equals I sq ft of external pipe surface. 
Hence, the lineal feet of I J in- pipe required is 

966 x Z.3 - 2220 ft 

11*21- Forced Convection Evaporator*. 

Forced convection evaporators, commonly 
called "unit coolers" or "blower coils" in com- 
mercial refrigeration, are essentially finned coils 
in a metal housing and equipped with 
or more fans to provide air circulation. 
Some typical unit cookrs on shown in Fig. 
11-30. 

The total cooling capacity of any evaporator 
is directly related to the air quantity (cfm) 
circulated over the evaporator. The air quantity 
required for a given evaporator capacity 
is basically a function of two factors: (i) the 
sensible heat ratio and (2} the drop in the 
temperature of the air passing over the cvapo* 
rator. vix: 

Cfm - 

Total capacity (Btu/hr) x sensible heat ratio 
Temperature drop of the air x 1.08 

(11-5) 

The sensible heat ratio is (he ratio of the 
sensible cooling capacity of the evaporator to 
the total cooling capacity. When air is cooled 



below its dew point temperature, both the tem- 
perature and the moisture content of the air an 
reduced (Chapter 5). The temperature reduc- 
tion is the result of sensible cooling, whereas the 
moisture removed is the result of latent coo ting. 
Kence. for an evaporator having a total cooling 
capacity of one ton (12,000 Btu/hr) and a 
sensible heat ratio of 0.85. the sensible cooling 
capacity of (he evaporator is 10,200 Btu/hr 
<W% or 12,000). whereas the latent cooling 
capacity is 1800 Btu/hr (12,000 - 10.200). 
Naturally, the sensible heal ratio of any 
evaporator will depend upon the conditions of 
the application, the design of the evaporate r, and 
the air quantity. An average sensible heat ratio 
for unit coolers is approximately 0.83. 

As a general rule, the air temperature drop 
through a well -designed unit cooler is approxi- 
mately one- half the difference between the space 
temperature and the refrigerant temperature. 
For example, for an evaporator TD of 10 F. 
the air temperature drop through the unit 
u.H)lcr should be .tppri^inuid', f I 

The constant LOS in Equation 1 1-3 is a con- 
version factor involving air density (0J5), air 
specific heat (0.24 Btu/lb/° F), and minutes per 
hour (*0).* 

£x m en p I b 1 1 -5. 1 V (ermine the approximate 
quantity of air {cfm) circulated over a unit 
cooler having a capacity of 20.000 Btu/hr if the 
sensible heat ratio is 0,85 and the design evapo- 
rator TD is 1 3* F. 

Solution. Apply- _ 30.000 Btu/hr x 0.83 
ing Equation 1 1-5, 7.5 x 1.08 

the air quantity - 2)00 cfm 

Ai a general rule, air velocities across the face 
of unit coolers are maintained between 300 and 
• Sen Section 14-4. 



186 PRINCIPLES OF REFRIGERATION 




Flf. Il-M. Topical Link tooter deslgni. Noikl thlt eool*r d*ligni iri mch that the air It not dttcharjid 
directly on i ho *to red product. (Counety Dunham-Bush, Inc.) 



500 ft per minute (Tpm), Although higher 
velocities will result in higher transfer coeffi- 
cients, they are not usually practical since they 
alio increase fan horsepower requirements. 
When the fan horsepower is increased beyond a 
certain point, the additional heal given off by 
the fan motor resulting from the increase it) 
horsepower will exceed the increase in unit 



cooler capacity resulting from higher air 
velocities. Hence, the net effect in such cases 
it to decrease, rather tkm increase, the over- 
all capacity of the unit cooler. Too, where 
the air velocity exceeds 500 fpm there Is a 
tendency for moisture to he blown from the 
face of the coil into the space and onto the 
product. 



The iir velocity {fpm) over the evaporator ii 
a function of the air quantity (din) and the face 
ana of the evaporator (v\ ft), viz : 

Air quantity (cfm) 



Velocity (fptro - 



(11-6) 



Face area |sq f i) 

Example 1 1 ■*. Determine the face am of 
thcevaporaior in Example 1 1 -5 if i he face velocity 
is to be maintained at J 50 fpm. 

Solution. By rearranging and _ 2100 dm 
applying Equation 11*6, the " 350 fpm 
face ana — 6 iq ft 

1 1-22- Rat i ng and Saf action of U n i t Coolers , 
Basic rating! for unil cooler* are given in Blu/hr/ 
* F TD. For convenience, sometimes ratings 
are given for 10" F and li* F TDs. Aa in the 
case Of natural convection evaporators, the 
deafen TD for unit cooler* depends primarily 
on the space humidity requirements. In general, 
for any given space humidity, the design TD for 
unit coolers is about 4* F lo ** F leas than those 
used for natural convection evaporator* (we 
Table 1 1-2). 

Since the air quantity is usually fixed by (he 
fan selection at the time of manufacture, realiz- 
ation of the rated capacity will depend primarily 
upon whether or not the coil is kepi reasonably 
free of frost by adequate defrosting When the 
space is maintained below 34' F, some means of 
automatic defrosting musi be used (see Chapter 
20). 



EVArOWKTORS 117 

Example I J -7. From Table R-8, select a 
forced convection evaporator (unit cooler) 
suitable for installation in a beer storage cooler 
having a calculated heal load of 16,600 Btu/hr. 
Since apace humidity is not a factor, use JO F 
TD for high system efficiency. 

Solution. From Table R-S, select unit cooler 
Model #UO!80 having a capacity of 18,000 
Btu/hr at a 10* F TD, Since the unit cooler fan 
motor operates inside the refrigerated space, the 
motor heat becomes a part of the space cooling 
load and must be added to the load calculations. 
From Table R-g, the heat given off by the Tan 
motor is 24,000 Btu/24 hr, Since the fan oper- 
ates continuously to provide air circulation in 
the refrigerated space, while the average hourly 
cooling load is based on a 16-hr running time, 
the average Btu per hour load resulting from 
the fan motor heat is 



34,000 BtiV?4 hr 
16 far 



- 1 300 Btu/hr 



Hence, when the fan motor beat is considered, 
the average hourly cooling load for the beer 
cooler becomes 18,100 Btu/hr (16,600 + 1500). 
Since the unit cooler selected has ■ capacity of 
IS.OOO Btu/hr, it will be adequate for the 
application. 

Suggested locations for unit coolers in walk- 
in refrigerators are shown in Fig. 11-11, 



ftg. IIO I. Su(|«ikwii for 
location of unit coolin In 
walk-in f*frS|tf*ton. (From 
tti* Mft£ Dots ftcok, D**Jjn 
Volume, 1957-M edition, re- 
produced by ptrmmlon of Ux 
American Society of Hmlnf. 
RvtriftrMint; and Air-Condi* 

tlMldf Enjin«rv) 




IBS PRINCIPLES Of REFRIGERATION 

11-13. Liquid Chilling Evaporators, As with 
air-cooling evaporators, liquid chilling evapo- 
rators vary in type and design according to the 
type of duly for which they art intended. Five 
general types of liquid chiller* are in common 
use: (1) the double^pipc cooler. (2) the Baudelot 
cooler . (3) the lank-type cooler. (4) theshell*and- 
coil cooler, and (3) the shell-and-tube cooler. 
In all cases, the factors which influence the 
performance of liquid chillers are the same as 
those which govern the performance of air- 
cooling evaporators and all other heat transfer 



headers and art connected together by re- 
movable return bends (inset I, The advantages 
claimed for ink unit are rigid construction, the 
elimination of refrigerant [uints, and easy 
accessibility of the inner tubes for cleaning. 

Double-pipe coolers ma;, ix; operated cither 
dry-expansion or floodci!. In either cue, 
counterflowing of the fluids in the lubes pro* 
duces a relatively high heai transfer coefficient. 
However, ihis type of cooler hits the disadvan- 
tage of requiring more space, particularly head 
room, than some of the oiher cooler design*. 



Current conservative design values of heat transfer coefficient C based on 
outside surface for bare tube coolers, unless mentioned otherwise, are as 
follows: 

Flooded shell -and- tube cooler (water to ammonia or R-12) 

Flooded ihell-and -finned tube high velocity R-12 water cooler 

Flooded shell-and-tube cooler (brine to ammonia) 

Flooded shell-and-tube cooler (brine to R-I2> 

Dry-expansion shel)-«nd-tube cooler, R-12 in tubes, water in shell 

Baudelot cooler. Hooded (ammonia or R-12 to water) 

Bau deloi cooler, dry-expansion (ammonia to water) 

Baudelot cooler, dry-expansion (R-12 to water) 

Double-pipe cooler (water to ammonia) 

Double-pipe cooler (brine to ammonia) 

Shell -and -coil cooler (water to ammonia) 

Shell-and-coil cooler (water to R-12) 

Spray-type shell-and-tube water coolers (ammonia or R-12) 

Tank-and-agitator, coil-type water cooler, ammonia, flooded 

Tank-and-agiiator, coil-type waler cooler. R-12 Hooded 

Tank, ammonia, brine cooling, coils between can in k* tank 

Tank, high velocity raceway type, brine to ammonia 

Fig. It-XL Ht« tniwftr coeftidants (or vsrfout types of liquid chlllan. (Rtpi-miad from JMJ-5* 
ASflE Dots Book, by furmlulon of th* Amtfican SoclKy of H**tlnj. ft«lri{«rsttr-i, «nd AJr-CondltJoiUnt 
En|in*tn,) 

For this reason, the double-pipe cooler is used 
only in some few special appliea lions. A number 
have been used in the wine-nuking and brewing 
industries to chill wine and wort, and in ihe 
petroleum industry for the shilling of oils. 
1 1 -IS. Baudelot Coolers- The Baudelot cooler 
shown in Hg. I l-M contisu of a series of hori- 
zontal pipes which are located One under the 
□I her and are connected i^geiher to form a 
refrigerant circuit or circuit*. For either dry- 
expansion or flooded operation, the refrigerant 
is circulated through the inside of the tubes 
while the chilled liquid flows in a thin film over 



Min 


|ft« 


su 


150 


JO 


150 


45 


100 


j« 


90 


50 


115 


100 


200 


60 


150 


60 


120 


50 


150 


50 


125 


10 


25 


10 


25 


ISO 


250 


so 


125 


60 


100 


15 


40 


SO 


110 



surfaces. Heat transfer coefficients for average 
designs of tome of the various chiller types are 
listed in the table in Fig. 1 1-32. 
11-24. Double-Pipe Coolers, The double-pipe 
cooler consists of two lubes SO arranged that 
one tube is inside the other. The chilled fluid 
flows in one direction through the inner tube 
while the refrigerant flows in the opposite 
direction through the annular space between the 
inner and outer tubes. One design of a double- 
pipe cookr is shown in Fig, 1 1-33. In this design 
the outer tubes arc welded to vertical refrigerant 
headers while ihe inner lubes pass through the 






EVAPORATORS IB? 





Fij. 11-11. Double pip* cootar. Removable mum b*ndi (right) ira ditlfnad to make tub* readily 
accessible, for clunlnj (Courtey Viltar rltnuficturini Company.) 



the outside. The liquid fowl down over the 
tubes by gravity from a distributor foaled at 
the top of the cooler and is collected in a trough 
at the bottom. The fact thai the chilled liquid 
it at atmospheric pressure and is open to the air 
nukes the Baudelot cooler ideal for any liquid 
chilling application when Aeration u a factor . 
The Baudejm chiller has been widely used for the 
cooling of milk, wine, and wort, and for the 



chilling of water Tor carbonation in bottling 
plan is. With this particular type of chiller it is 
possible lo chili liquid to a temperature very 
close to the freezing point without the danger 
of damaging the equipment if occasional 
freezing of the liquid occurs. 

Another advantage of the Baudelot cooler, 
and one which is shared by the double-pipe 
cooler, is that the refrigerant circuit is readily 



Flf- 1 1-14. Bmddoi tooltr 
employ *d in milk-cool I nj 
application. (Coortu/ Dola 
Refriierailn g Company.) 




190 PRINCIPLES OF REFRIGERATION 
_ Warm fcquWI m 




Flf. If-JS. Typlal construc- 
tion of tnnk-cjfpe liquid Caol a r. 



J?Bfnj(er*n1 

fines 



split into several parts, a circumstance which 
permits precoohng of the chilled liquid with 
cold wain: before the liquid enter* the direct- 
expansion portion of the cooler (see Fig. 1 7-34>. 
11-16. Tank-Type Cooler*. The lank-type 
liquid chiller consuls essentially of A bane-lube 
refrigerant coil installed in the center or si one 
side of a large steel lank which contain} the 
chilled liquid. Although! completely Autxnefgctl 
in the chilled liquid, the refrigerant coil is 
separated from the main body of the liquid by 



a baffle arrangement. As shown in Fig. 11-35. 
a motor driven agitator is unlived to circulate 
the chilled liquid over the cooling coil at 
relatively high velocity, usual h, between LOO and 
1 50 ft per minute, the liquet being drawn in at 
one end of the coil compart men I .ind discharged 
at the other cod. 

The spiral-shaped, bare-mbc coils mentioned 
in Section 1 1*4 and the race t-jt -type coil illus- 
trated in Fig. 1 1-36 arc two coO designs fre- 
quently employed in tank-i>pe chillers. With 




M. Flooded rac*w*f colt, (Courttiy Vllltr Hlftufatlu t\ n| Company,} 



IVAPOMATOHS 



191 



either design the coils are operated flooded. 
The lee-Cel shown in Fig. 11-11 is another 
variation or the tank-type chiller. 

Tank -type chiller* can be Applied to any 
liquid-chilling application where sanitation is 
not a primary factor, and arc widely used for 
the chilling or water, brine, and other liquids 
to be used as secondary refrigerants. Because 



a welded steel shell (Fig. 1 1-37), As a genera] 
rule, the chiller is operated dry -expansion with 
the refrigerant in the coils and ihe chilled liquid 
imhc*hell. In a few cases, the chiller is operated 
flooded, in which esse the refrigerant is in the 
shell and the chilled liquid passe* through the 
tubes. The former arrangement has ihe ad van* 
lage of providing a holdover capacity, thereby 



liquxs line 



ntVffMlM 



tc uuen dfatn 



Thtrmal eaptAiion vahw 




Flj. I I.JT- Shrll-ind-cdl cootsr, {Courtis/ Acm* tndintriai.) 



of their inherent holdover capacity, they are 
particularly suitable for applications subject to 
frequent and seven fluctuations in loading. In 
such cases, a comparatively large chilled-liquid 
storage tank is provided in order to minimize 
the rise in the temperature of the chilled liquid 
during periods of peak demand. The advantage 
gained by precooling is often considerable in 
cases where the liquid to be chilled enters the 
cooler at relatively high temperatures. 
11-27. She 1 1 -and -Coil Cooler*. The shell- 
and-coil chiller is usually made up of one or 
more spiral-shaped, bare-tube coils enclosed in 



making this type of chiller ideal for small 
applications having high but infrequent peak 
toads, tl is used primarily for the chilling of 
water for drinking and for other purposes where 
sanitation is a prime factor, as in bakeries and 
photographic laboratories. 

When operated Hooded with the refrigerant 
in the shell, this type of chiller becomes what 
is commonly referred to as an "instantaneous" 
liquid chiller. One of the disadvantages of this 
arrangement is that there is no holdover 
capacity. Since the liquid is not recirculated, 
it must be chilled instantaneously as it passes 



192 PRINCIPLES OF REFRIGERATION 



Liquid Outlet - 




MM *c r ^"t , £ t i , 'l 



f*l 



Bf> tl-3*. Typical iheilmd-tub* chilltn. (0) Hooded typ«- Tuba bundla li ramovablc |b) Dry-txpuuion 
Eype (refrigerant in lubm). Not* taffllnf of Wiur circuit Tub* theati ire fixed. (C*ur(t*y WonMnftOII 



ihrough the coils. Another disadvantage is that 
Ihc danger of damaging the chiller in (he event 
of freeze-up is greatly increased in any chiller 
where the chilled liquid is circulated through 
the coils or tubes, rather than over the outside of 
the tubes. For this reason, chillers employing 
this arrangement cannot be recommended for 
any application where it is required to chill the 
liquid below 38° F, 

Instantaneous shell -and -coil chillers arc used 
principally for the chilling of beer and oiher 
beverages in "draw-bars," in which case ihc 
beverage is usually precooled to some extent 
before entering the chiller. 
1 1-2& Shell-and-Tvbe Chillers, Shell-and- 
tube chillers have a relatively high efficiency, 
require a minimum of floor space nnd head 
room, are easily maintained, and are readily 
adaptable to almost any type of liquid-chilling 
application. For these reasons, the shcll-and- 



tube chiller it by far the mosi widely used type. 
Although individual designs il filer somewhat, 
depending upon Ihc refrigen-ini used and upon 
whether the chiller is operated dry-expansion 
or flooded, the shcli-and-tube chiller consists 
essentially of ft cylindrical Keel shell in which a 
number of Straight tubes are arranged in parallel 
and held in place at the ends by tube sheets. 
When the chiller is operated dry-expansion, the 
refrigerant is expanded into the tubes while the 
chilled liquid is circulated ilmmgh the shell 
(Fig. 1 1 -38b). When the chiller is operated 
flooded, the chilled liquid is circulated through 
the 1 iilics .111 J the refrigerant ii contained in the 
shell . the level of the liquid refrigerant in the 
shell being maintained with some type of float 
control {Fig. LlOBd), In either case, the chilled 
liquid is circulated through the chiller and con- 
necting piping by means of a liquid circulating 
pump, usually of the centrifugal type. 



EVAPORATORS t?J 



Shell diameters for shell~and-tube chillers 
range from approximately 6 to GO in,, and the 
number of tubes in the shell varies from fewer 
Lhan SO to several thousand- Tube dkmclem 
range from I in. through 2 in., and lube lengths 
vary from S to 20 ft. Chillers designed for use 
with ammonia are equipped with steel tubes, 
whereas those intended for use with other 
refrigerant! an usually equipped with copper 
tubes in order to obtain a higher heat transfer 
ooefnaenl. Because of the rela lively low film 
conductance of halocarbon refrigerants, chillers 
designed for use with these refrigerants are often 
equipped with tubes which are fumed on the 
refrigerant side. In the case of dry-expansion 
chillers, the lubes are finned internally with 
longitudinal fins of the types shown in Fig. 
11-12. For flooded operation, the tubes are 
finned externally using a very short fin which 
protrude* from the lube wall only approximately 
,\thof an inch. 

As a general rule, dry<*p*nlkm drillers are 
employed in small and medium tonnage instal- 
lations requiring capacities ranging from 2 to 
approximately 250 tons, but ore available in 
larger capacities. Flooded chillers, available 
in capacities ranging from approximately 10 
through several thousand tons, are more 
frequently applied in ihe larger tonnage instal- 
lations. 

Il-M. Dry-Expansion CMEIar*. The principal 
advantages of the dry-expansion chiller over the 



flooded type me the smaller refrigerant charge 
required and the assurance of positive oil return 
to the compressor. Too, aa previously stated, 
the possibility of damage to the chiller in the 
event of freeze-up is always considerably less 
when the chilled liquid is circulated over the 
lubes ml her lhan through them. The more 
important construction details of several designs 
of dry-cupansion chillers are shown in Figs. 
1 1-39 "and 1 1-40. 

En order to maintain the liquid velocity 
wilhin ihc limits which will produce the most 
effective heat transfer-pressure drop ratio, ihe 
vctochy of the chilled liquid circulated over the 
lubes is controlled by varying the length and 
spacing of the segmental baffles, When the 
fiow rate and/or liquid viscosity is high, short, 
widely spaced taffies are used to reduce the 
velocity and minimize the pressure drop through 
the chiller. When the flow rate and/or liquid 
viscosity is low. longer, more closely spaced 
baffles are used in order to increase fluid 
velocity and improve the heat transfer coef- 
ficient (Fig. IMkr). 

The number and the length of the refrigerant 
circuits required to maintain the refrigerant 
velocity through the chiller tubes within reason- 
able limits depend on the total chiller load and 
on the relationship of the chilled liquid flow 
rate to the METD. Since the* factors vary 
with the individual application, it follows that 
the optimum refrigerant circuit design atso 




Flf . I !-«. C u«w*¥ Mrtloit llltMirstini eoimrucrion dtutli of drr-wtpUMlon chllfer with ftxad tub* ihMta, 
(Counter Aem* Industrial.) 



SM PRINCIPLES OF REFRIGERATION 




Flf. I MO. Dry-expirtKon (hilJ*r with tub* bundl* fnnl*riy rtmovtd IO llMW lutw inwjimint mnif 
refrlferani flUtribirtcn, Tub* bundl* w b* rtrnov*d U t unll, (CourtMjf Ktnmrd Cn.non, Amtritln Air 
Fihmr Com piny , Inc.) 



varies with the individual application. For this 
reason, chillers arc made available with cither 
single or multiple refrigerant circuits of varying 
lengths. For the design shown in Tig. 11-39, 
the number and length or the refrigerant circuits 
depend on the lube length and on the arrange- 
ment of the baffling in the end-plates or refriger- 
ant beads which are bolted to the lube sheets 
at the ends of the chilter. The refrigeranl 
circuit arrangement for any one model chiller 
can be changed by -changing the refrigerant 
heads (Fig. 11 -41 A}' 

11-30. Flooded Chilian, Standard flooded 
chiller designs include both single and multipass 
lube arrangements, lor single pass flow, the 
tubes ere so arranged lhat ihe chilled liquid 
passes through all the lubes simultaneously 
and in only one direction. 

Multipass circulation of the chilled liquid 
through the ehilkr is accomplished through the 
use of baffled end-plates or heads which are 
boiled lo the ends or the chiller (Fig. 1 1-42), 
The arrangement of the end-plate baffling 
determines the number of passes the chilled 
liquid makes from one to the other before 
leaving the chiller. Although two, four, and 



siX'put arrangements are ihe most common, 
more pears are used in many instances. 

As in the case of the dr\ -expansion chiller, 
some Hooded chillers are deu^ned with remov- 
able tube bundles, whereas other? have fixed 
lube sheets. In the fixed tube sheet design, the 
tube sheets are welded to the shell so that 
the lube bundle is not removable. However, 
by unbolting the end-plate* the tubes become 
readily accessible for cleaning and individual 
lubes can be removed and replaced if necessary. 
The chillers shown in Fig* LI 3Bfr and 11-39 
employ fixed tube sheets, whereas those in Figs. 
M->3Bu and I MO have tube huiidlcs. 

In some flooded chiller designs, the shell 
is only partially filled with tubea in order to 
provide a large vapor-disen^igmg area and rela- 
tively low velocity in the spate above the tubes 
(Fig. 11*43). Th n design dim i n.j | e<i the possibility 
of liquid carry-over into the suction line and 
therefore is particularly well mi i fed to sudden 
heavy increases in loading 

In those chiller designs where Ihe shell is 
completely filled with tubes, a surge drum or 
accumulator should be used in separate any 
entrained liquid from the vapor before the vapor 






EVAPORATORS IW 



enter* the suction line. Some flooded chiller* 
■re equipped with bull [-in liquid-suction heal 
exchanger* (Fig. 11-45). Although I he prima ry 
function of the heal exchanger is to insure that 
oniy dry vapor enter* the suction line, it has 
the additional effect of increasing the efficiency 
of the chiller in that it subcools the liquid 
approaching the chiller and thereby reduces the 
amount of flash gas that enter* the cooler. 




Short cut on-*** ipKir^g 




taint* tm n**cs -intern*! vww 



4-FHs 

JOtiHh 




a-ataa 

2 Circuits 



•^ 





Flf. I Ml. (o) Baffla ipiCinf in d ry-** pwulon ehil- 
kfi. (D) Typ+oJ refrliertnt headl far dry-*Kp*nile>n 
chiller. (CDurte»r of Acme InduitrkM.) 



The vertical ihcll-and-lubc chiller shown in 
Fig. 11-44 has the advantage of requiring a 
minimum amount of floor space. The chiller 
is operated flooded. The chilled liquid enters 
the chiller at the top and flows by gravity down 
the inside of the tubes. A circulating pump 
draws chilled liquid from the storage tank at the 
bottom and delivers it through the connecting 
piping- The return liquid is piped to the dis- 
tributor box at the top. from where it again 
flown through the lubes. A specially designed 
distributor installed at the lop of each tube 
(inset) imparts a. swirling motion to the chilled 
liquid, which causes the liquid to flow in a. com- 
paratively thin film down the inside tube 
turfites, 

11-31. Spray-Type Chillers. The spray-type 
chiller is similar in construction to the conven- 
tional flooded chiller cuoep! that the liquid 
refrigerant is sprayed over the outside of the 
waier tubes from nozzle* located in a spray 
header above the tube bundle (Fig. 20-19). The 
unevaporated liquid drains from the tube into a 
tump at the bottom of the chiller from where it 
is recirculated to the spray nozzles by ■ low head 
liquid pump. A high recirculation rate assures 
continuous wetting of the tube surfaces and 
results in a high rale of heat transfer. 

The principal advantages of this type of 
chiller is its high cflicicncy and relatively small 
refrigerant charge. Disadvantages are the high 
installation cost and the need for a liquid 
recirculating pump. 

1 1 ■&. Chiller Selection Procedure. Although 
selection methods differ somewhat depending 
upon the type of chiller and the particular manu- 
facturer, all are based on the simple funda- 
mentals of heat transfer and fluid now which 
have already been described. Almost without 
exception, manufacturers include sample selec- 
tion procedure along with the design and 
capacity data in their e quip ment catalogs. The 
following selection procedure follows very 
closely that given in the catalog of one 
manufacturer for the selection of dry-expansion 
chillers-* 

Example II-*, ]i is desired to cool 50 gpm 
of wrner from 54° F to 46° F with a refrigemnt 
tempera lure ai measured at the cooler outlet of 
40- F usinj; Refrigerant* 1 2. 

* Acme Industrie*. Inc. 



W PRINCIPLES OF REFRIGERATION 




Fig. 11*41. Flooded thlller d**F(n«d far multlp»I circulation of chiliad liquid. I",>. pm circulation i* 
ttcompliihed by mean* of th* C4ffl«d *fld-£l«*-, or writer heads which ir* bc-lud to Ihe endi of the chiller. 

{CourtHr Vilter Minu^ctuhni Company.) 



Solution 

Step 1. Determine ihe total chiller load 



in 



tons. 



produces the highest pumping had Hence, if 
space it not ■ problem, ihe must logical choice 
would lean to be type 8M I U-wcvcr, a check 



Gpm x 500 x cooling range 

1 2.000 Blu/hr/ton 
SO x 500 x (54 - 46) 



> 2,000 



— 16.7 ions 



Step 2. Determine the mean effective tem- 
perature difference (METD), 
Water in minus refrig- 
erant temperature 54 — 40 ■■ 14* F LTD 
Water out minus re- 
frigerant temperature -Mi - 40 — 6" F STD 
From Table II -1, METD - 9.47* F 

Step 3. Select trial chiller {shell diameter and 
baffles spacing) from Fig, I of Table R-9. Enter 
Fig. 1 at 50 gpm on the lower vertical scale and 
move horizontally across the chart to the 
diagonal line representing the type unit desired, 
The number indicates shell diameter and the 
letter indicates baffle spicing, Possible choices 
are 10M, I2L, SM, 12K, 10K, and SL, As a 
general rule, small diameter chillers are more 
economical, whereas large diameter chillers are 
more compact. Type M burning produces the 
lowest pumping head, whereas type L baffling 




Fig. I ML Flooded chiller with shell only partially 
filled with tubst In order to provide i lirgc vapor 
d-Hniafln & tree ibovi the tubri. {CourtMy Worth- 
ln|tdn Corporation.) 



EVAPORATORS 



197 



will show that neither 8M nor 8L is available 
with sufficient surface area in this instance. 
Therefore, select type 10M (8 to 30 tons). From 
the point of intersection move vertically upward 
to a diagonal line in the upper portion of Fig. 1 
which represents a METD of 9.47° F as found 



Water 
inlet 



Gas outlet - 



14 ft DXH chiller has a surface area of 184 sq ft 
(Model No. DXH-1014). 

Step 6. Determine the water pressure drop 
through the chiller. From the bottom of Fig. 1, 
Table R-9, the pressure drop per foot of length 
with type M baffling is 0.425 ft of water column. 



Refrigerant 
feed 





Inset 



Fig. 11-44. Vertical shell-and-tube "Spira-Flo" chiller designed for flooded operation. The water flowing 
down through the tubes is given a swirling action by specially designed nozzles (inset). (Courtesy Worthington 
Corporation.) 



in Step 2. From this intersection move hori- 
zontally to the scale at the left margin and read 
the loading of 1110 Btu/hr/sq ft (loading is the 
U value times the METD). 
Step 4. Determine the surface area required. 

Surface area -^^ 
Loading 

200.000 

-^5-- 180.2 sq ft 

Step 5. Select chiller length from Table R-9 
to meet surface area requirements. A 10 in. x 



Pressure drop = length (feet) x pressure drop/ 
foot 

14 feet x 0.425 = 5.95 ft H a O 

11-33. Direct and Indirect Systems. Any 

heat transfer surface into which a volatile liquid 
(refrigerant) is expanded and evaporated in 
order to produce a cooling effect is called a 
"direct-expansion" evaporator and the liquid 
so evaporated is called a "direct-expansion" 
refrigerant. A direct-expansion or "direct" 



198 PRINCIPLES OF REFRIGERATION 



Brine coil "V. 



WWW/WW/WW/W/W, 



Refrigerated 



c 



space or 



material 



V/s;///;ss///s;;j;ss;;ss;;;//v. 



Liquid from 
receiver 



from /" 
ver jfVl 

* — &-i 



Vapor to 

compressor 

suction 



Warm brine_ 
to chiller " 



Refrigerant control 



I Brine solution! 



?mrm/m/m///??r#mrA 




Cold brine 
to coil 



Brine pump 



Fig. I M5o. Indirect system. 

■ Brine coil 



t4- 



Cold air to 
refrigerated space 



Warm brine 
to chiller 



receiver ,^^BZ^mB^. 




Refrigerant control 



'**r . 



Vapor to rx. s; 

compressor z | 

suction vM 



.. . 

^ £ " w wiw hw i p wbh w/ ■* >7 



Brine solution { 






7~\ 



Cold brine 
to coil 




Warm air 
from space 




-Air duct 



^ ^-Brine 



-Brine pump 
Fig. Il-45b. Indirect system — brine coil in communicating duct. 



refrigerating system is one wherein the system 
evaporator, employing a direct-expansion refrig- 
erant, is in direct contact with the space or 
material being refrigerated, or is located in air 
ducts communicating with such spaces. Up to 
this point, only direct refrigerating systems have 
been considered. 

Very often it is either inconvenient or un- 
economical to circulate a direct-expansion 
refrigerant to the area or areas where the cooling 



is required. In such cases, an indirect refrigerat- 
ing system is employed. Water or brine (or 
some other suitable liquid) is chilled by a direct- 
expansion refrigerant in a liquid chiller and 
then pumped through appropriate piping to the 
space or product being refrigerated. The 
chilled liquid, called a secondary refrigerant, 
may be circulated directly around the refriger- 
ated product or vessel or it may be passed 
through an air-cooling coil or some other type 



EVAPORATORS 



199 



of heat transfer surface (Fig. 11-45). In either 
case, the secondary refrigerant, warmed by the 
absorption of heat from the refrigerated space 
or product, is returned to the chiller to be 
chilled and recirculated. 

Indirect refrigerating systems are usually 
employed to an advantage in any installation 
where the space or product to be cooled is 
located a considerable distance from the con- 
densing equipment. The reason for this is that 
long direct-expansion refrigerant lines are 
seldom practical. In the first place, they are 
expensive to install and they necessitate a large 
refrigerant charge. Too, long refrigerant lines, 
particularly long risers, create oil return prob- 
lems and cause excessive refrigerant pressure 
losses which tend to reauce the capacity and 
efficiency of the system. Furthermore, leaks 
are more serious and are much more likely to 
occur in refrigerant piping than in water or 
brine piping. 

Indirect refrigeration is required also in many 
industrial process cooling applications where 
it is often impractical to maintain a vapor tight 
seal around the product or vessel being cooled. 
Too, indirect systems are used to an advantage 
in any application where the leakage of refriger- 
ant and/or oil from the lines may cause con- 
tamination or other damage to a stored product. 
The latter applies particularly to meat packing 
plants and large cold storage applications when 
ammonia is used as a refrigerant. 
1 1-34. Secondary Refrigerants. Some com- 
monly used secondary refrigerants are water, 
calcium chloride and sodium chloride brines, 
ethylene and propylene glycols, Methanol 
(methyl alcohol), and glycerin. 

Almost without exception, water is used as 
the secondary refrigerant in large air condition- 
ing systems and also in industrial process cooling 
installations where the temperatures maintained 
are above the freezing point of water. Water, 
because of its fluidity, high specific heat value, 
and high film coefficient, is an excellent 
secondary refrigerant. It also has the advantage 
of being inexpensive and relatively noncorrosive. 
In air conditioning applications, the chilled 
water is circulated through an air cooling coil 
or through a water spray unit. In either case, 
the air is both cooled and dehumidified. In the 
water spray unit, the water i. sprayed from 
nozzles and collected in a pan or basin at the 



bottom of the spray unit, from where it is 
returned to the chiller. Since the air passing 
through the water spray is chilled below its dew 
point temperature, a certain amount of water 
vapor is condensed from the air and is carried 
to the basin with the spray water. With either 
the cooling coil or the spray unit, the amount 
of cooling and dehumidification can be con- 
trolled by varying the amount and temperature 
of the chilled water. 

Water is also used frequently as a secondary 
refrigerant in small beverage coolers and in 
farm coolers designed for cooling milk cans. 
In such cases, the water, because of its high 
conductivity, permits more rapid chilling of the 
product than would be possible with air. Too, 
the water supplies a holdover capacity which 
tends to level out load fluctuations resulting 
from intermittent loading of the cooler. 
11-35. Brines. Obviously, water cannot be 
employed as a secondary refrigerant in any 
application where the temperature to be main- 
tained is below the freezing point of water. In 
such cases, a brine solution is often used. 

Brine is the name given to the solution which 
results when various salts are dissolved in water. 
If a salt is dissolved in water, the freezing tem- 
perature of the resulting brine will be below the 
freezing temperature of pure water. Up to a 
certain point, the more salt dissolved in the 
solution, the lower will be the freezing tempera- 
ture of the brine. However, if the salt concen- 
tration is increased beyond a certain point, the 
freezing temperature of the brine will be raised 
rather than lowered. Hence, a solution of any 
salt in water has a certain concentration at 
which the freezing point of the solution is lowest. 
A solution at the critical concentration is called 
a eutectic solution. At any concentration above 
or below this critical concentration, the freezing 
temperature of the solution will be higher, that 
is, above the eutectic temperature.* When the 
salt content of the brine is less than that which 
is required for a eutectic solution, the excess 
water will begin to precipitate from the solution 
in the form of ice crystals at some temperature 
above the eutectic temperature. The exact tem- 
perature at which the ice crystals will begin to 

* At any temperature other than the eutectic 
temperature, the term "freezing temperature" is 
used to mean the temperature at which ice or salt 
crystals begin to precipitate from the solution. 



200 PRINCIPLES OF REFRIGERATION 



form depends upon the degree of the salt con- 
centration and upon the relative solubility of the 
salt in water, the latter factor decreasing as 
the temperature of the solution decreases. The 
continued precipitation of ice crystals from the 
solution as the temperature is reduced causes a 
progressive increase in the concentration of the 
remaining brine until, at the eutectic tempera- 
ture, a slush consisting of ice and eutectic brine 
will exist. The further removal of heat from this 
mixture will result in solidification of the eutec- 
tic brine. Solidification of the eutectic brine will 
take place at a constant temperature. 

On the other hand, when the salt content of 
the brine is in excess of the amount required for 
a eutectic solution, the excess salt will begin to 
precipitate from the solution in the form of salt 
crystals at some temperature above the eutectic 
temperature. Continued precipitation of salt 
from the mixture as the temperature is reduced 
will result in a mixture of salt and eutectic brine 
when the eutectic temperature is reached. The 
further removal of heat from the mixture will 
result in solidification of the eutectic brine at 
constant temperature. 

Two types of brine are commonly used in 
refrigeration practice: (1) calcium chloride and 



Air - 




Air 



iff Mr M 



1r 



m 



ii '" 



i 



Brine from 
chiller 



I 



Jl 



Brine to 
concentrator 



Concentrator 




I Brine 



\, Brine from 
concentrator 

*" Brine to 
chiller 



Fig. 1 1-46. Brine spray cooler. 



(2) sodium chloride. The two brines are pre- 
pared from calcium chloride (CaCl^ and sodium 
chloride (NaCl) salts, respectively, the latter 
salt being the common table variety. 

Calcium chloride brine is used primarily in 
industrial process cooling, in product freezing 
and storage, and in other brine applications 
where temperatures below 0° F are required. 
The lowest freezing temperature which can be 
obtained with calcium chloride brine (the eutec- 
tic temperature) is approximately —67° F. The 
salt concentration in the eutectic solution is 
approximately 30% by weight. The freezing 
temperature of various concentrations of cal- 
cium chloride brine are given in Table 11-3, 
along with some of the other important pro- 
perties of the brine. 

The principal disadvantage of calcium 
chloride brine is its dehydrating effect and its 
tendency to impart a bitter taste to food pro- 
ducts with which it comes in contact. For this 
reason, when calcium chloride brine is used in 
food freezing applications, the system must be 
designed so as to prevent the brine from coming 
into contact with the refrigerated product. 

Sodium chloride brine is employed mainly 
in those applications where the possibility of 
product contamination prevents the use of 
calcium chloride brine. Sodium chloride brine 
is employed extensively in installations where 
the chilling and freezing of meat, fish, and other 
products are accomplished by means of a brine 
spray or fog. 

The lowest temperature obtainable with 
sodium chloride brine is approximately —6° F. 
For this freezing temperature the salt concen- 
tration in the solution is approximately 23%. 
The thermal properties of sodium chloride brine 
at various concentrations is given in Table 11-4. 

It is of interest to notice that the thermal 
properties of both calcium chloride and sodium 
chloride brines are somewhat less satisfactory 
than those of water. As the salt content of the 
brines is increased, the fluidity, specific heat 
value, and thermal conductance of the brines all 
decrease. Hence, the stronger the brine solu- 
tion, the greater the quantity of brine that must 
be circulated in order to produce a given 
refrigerating effect. 

Since the specific gravity of the brine increases 
as the salt concentration increases, the degree of 
salt concentration and the thermal properties of 



EVAPORATORS 201 



Air -*■ 




Fig. 1 1-47. Brine spray cooler. Brine WW- 



Direct 
expansion ■ 
evaporator 



M Eliminators 1 



■■MhHAai.'i 




ut 



Brine to 
concentrator 



To condensing 
unit 



the brine can be determined by measuring the 
specific gravity of the brine with a hydrometer. 
11-36. Antifreeze Solutions. Certain water 
soluble compounds, generally described as anti- 
freeze agents, are often used to depress the 
freezing point of water. The more widely known 
antifreeze agents are ethylene glycol, propylene 
glycol, Methanol (methyl alcohol), and glycerin. 
All these compounds are soluble in water in all 
proportions. The freezing temperature of water 
in solution with various percentages of each of 
these compounds is given in Table 11-5. 

Propylene glycol is probably the most exten- 
sively used antifreeze agent in refrigeration 
service. In common with ethylene glycol, 
propylene glycol has a number of desirable 
properties. Unlike brine, glycol solutions are 
noncorrosive. They are also nonelectrolytic and 
therefore may be employed in systems containing 
dissimilar metals. Being extremely stable com- 
pounds, glycols will not evaporate under normal 
operating conditions. Because of the many 



advantages of glycol solutions, they are being 
used to replace brines in a number of installa- 
tions, particularly in the brewing and dairy 
industries. The change-over from brine to glycol 
can be accomplished with practically no change 
in the plant facilities. 

1 1-37. Brine Spray Units. Like chilled water, 
the chilled brine (or antifreeze solution) may be 
circulated directly around the refrigerated pro- 
duct or container, or it may be used to cool the 
air in a •refrigerated space. When used to cool 
air, the chilled brine is circulated through a 
serpentine coil or through a brine spray unit. 
Two types of brine spray units which have been 
used extensively are shown in Figs. 11-46 and 
11-47. In the former unit chilled brine from a 
brine chiller located outside the refrigerated 
space is- sprayed down from spray nozzles and 
collected in the basin of the unit, from where it 
is returned to the brine chiller. In the latter type 
the brine is chilled by a direct-expansion coil 
located within the brine spray unit itself. 



202 PRINCIPLES OF REFRIGERATION 

PROBLEMS 

1. A walk-in cooler 8 ft by 9 ft by 9 ft high has 
walls 6 in. thick and is maintained at a tempera- 
ture of 35° F. The load on the cooler is 7500 
Btu/hr. 

(a) Select a natural convection cooling coil 
(Plasti-Cooler) which will produce a 
relative humidity of approximately 85% 
in the cooler. 

(6) Select a Unit Cooler which will produce 
approximately the same conditions in the 
cooler. 

2. A freezing cabinet 6 ft high, 25 in. deep, and 
80 in. wide has a freezing load of 3600 Btu/hr. 



Based on an evaporator TD of 10° F, determine 
the size and number of individual freezer plates 
to be used as shelves in the freezing cabinet. 

3. The load on a tank-type brine cooler is 
4500 Btu/hr. The brine is to be maintained at 
a temperature of 35° F with a refrigerant tem- 
perature of 19° F. Assuming little or no 
agitation of the brine, determine the lineal feet 
of | in. pipe required for the evaporator. 

4. It is desired to cool 100 gpm of water from 
56° F to 46° F with a refrigerant temperature 
38° F at the cooler outlet. Select an appro- 
priate chiller and determine the water pressure 
drop through the chiller in psi. 



12 

Performance 
of Reciprocating 
Compressors 



12-1. Refrigeration Compressors. Vapor 
compressors used in refrigeration are of three 
principal types: (1) reciprocating, (2) rotary, 
and (3) centrifugal. Of the three, the recipro- 
cating compressor is by far the" one most 
frequently used. 

Rotary compressors are limited to use in very 
small fractional horsepower applications, such 
as home refrigerators and freezers and small 
commercial applications. Even in this limited 
area, rotary compressors represent only a small 
fraction of the total number used. Some rotary 
compressors are used also as booster compres- 
sors.* Their use for this purpose appears to be 
increasing. 

Centrifugal compressors are used only on 
very large applications, usually at least SO tons 
or above. In this area, they are widely accepted 
and are rapidly increasing in number because 
the number of large applications is growing 
steadily. 

Only the performance of reciprocating com- 
pressors will be discussed in this chapter. 
Reciprocating compressor design, along with 
the design and performance of rotary and centri- 
fugal compressors, is discussed elsewhere in the 
text at a more appropriate time and place. 
However, much that is said in this chapter about 



* Booster compressors are discussed in Chapter 



20. 



the performance of reciprocating compressors 
will apply also to the performance of rotary and 
centrifugal compressors. 
12-2. The Compression Cycle. Before at- 
tempting to analyze the performance of the 
compressor, it is necessary to become familiar 
with the series of processes which make up 
the compression cycle of a reciprocating com- 
pressor. 

A compressor, with the piston shown at four 
points in its travel in the cylinder, is illustrated 
in Fig. 12-1. As the piston moves downward 
on the suction stroke, low-pressure vapor from 
the suction line is drawn into the cylinder 
through the suction valves. On the upstroke of 
the piston, the low-pressure vapor is first com- 
pressed and then discharged as a high-pressure 
vapor through the discharge valves into the head 
of the compressor. 

To prevent the piston from striking the valve 
plate, all reciprocating compressors are designed 
with a small amount of clearance between the 
top of the piston and the valve plate when the 
piston is at the top of its stroke. The volume of 
this clearance space is called the clearance 
volume and is the volume of the cylinder when 
the piston is at top dead center. 

Not all the high-pressure vapor will pass out 
through the discharge valves at the end of the 
compression stroke. A certain amount will 
remain in the cylinder in the clearance space 
between the piston and the valve plate. The 
vapor which remains in the clearance space at 
the end of each discharge stroke is called the 
clearance vapor. 

Reference to Figs. 12-2 and 12-3 will help to 
clarify the operation of the compressor. Figure 
1 2-2 is a time-pressure diagram in which cylinder 
pressure is plotted against crank position. 
Figure 12-3 is a theoretical pressure- volume 
diagram of a typical compression cycle. The 
lettered points on theTP andPV diagrams corre- 
spond to the piston positions as shown in 
Fig. 12-1. 

At point A, the piston is at the top of its 
stroke, which is known as top dead center. When 
the piston is at this position, both the suction 
and discharge valves are closed. The high 
pressure of the vapor trapped in the clearance 
space acts upward on the suction valves and 
holds them closed against the pressure of the 
suction vapor in the suction line. Because the 



203 



204 PRINCIPLES OF REFRIGERATION 

pressure of the vapor in the head of the com- 
pressor is approximately the same as that of the 
vapor in the clearance volume, the discharge 
valves are held closed either by their own weight 
or by light spring loading. 

As the piston moves downward on the suction 
stroke, the high-pressure vapor trapped in the 
clearance space is allowed to expand. The 



Discharge^ 











180 
Crank position 



Fig. 12-2. Theoretical time-pressure diagram of 
compression cycle in which cylinder pressure is 
plotted against crank position. 



piston reaches the bottom of its stroke at point 
C. During the time that the piston is moving 
from B to C, the cylinder is filled with suction 
vapor and the pressure in the cylinder remains 
constant at the suction pressure. At point C, 
the suction valves close, usually by spring action, 
and the compression stroke begins. 

The pressure of the vapor in the cylinder 
increases along line C-D as the piston moves 
upward on the compression stroke. By the time 
the piston reaches point D, the pressure of the 
vapor in the cylinder has been increased until 
it is higher than the pressure of the vapor in the 
head of the compressor and the discharge valves 
are forced open; whereupon the high-pressure 
vapor passes from the cylinder into the hot gas 
line through the discharge valves. The flow of 



Fig. 12-1. (a) Piston at top dead center, (b) Suction 
valves open, (c) Piston at bottom dead center, (d) 
Discharge valves open. 



expansion takes place along line A-B so that 
the pressure in the cylinder decreases as the 
volume of the clearance vapor increases. When 
the piston reaches point B, the pressure of the 
re-expanded clearance vapor in the cylinder 
becomes slightly less than the pressure of the 
vapor in the suction line; whereupon the suction 
valves are forced open by the higher pressure in 
the suction line and vapor from the suction line 
flows into the cylinder. The flow of suction 
vapor into the cylinder begins when the suction 
valves open at point B and continues until the 




Clearance 



Volume of 

re-expanded • 

clearance vapor 



Volume 



Fig. 12-3. Pressure-volume diagram of typical com- 
pression cycle. 



PERFORMANCE OF RECIPROCATING COMPRESSORS 205 



the vapor through the discharge valves con- 
tinues as the piston moves from D to A while 
the pressure in the cylinder remains constant at 
the discharge pressure. When the piston returns 
to point A, the compression cycle is completed 
and the crankshaft of the compressor has 
rotated one complete revolution. 
12-3. Piston Displacement. The piston dis- 
placement of a reciprocating compressor is the 
total cylinder volume swept through by the 
piston in any certain time interval and is usually 
expressed in cubic feet per minute. For any 
single-acting, reciprocating compressor, the pis- 
ton displacement is computed as follows: 



v,-- 



n-Z)* x L x N x n 
4 x 1728 



(12-1) 



where V„ -» the piston displacement in cubic 

feet per minute 
D = the diameter of the cylinder (bore) 

in inches 
L = the length of stroke in inches 
N = revolutions of the crankshaft per 

minute (rpm) 
n — the number of cylinders 

The volume of the cylinder which is swept 
through by the piston each stroke (each revolu- 
tion of the crankshaft) is the difference between 
the volume of the cylinder when the piston is 
at the bottom of its stroke and the volume of 
the cylinder when the piston is at the top of its 
stroke. This part of the cylinder volume is 
found by multiplying the cross-sectional area of 
the bore by the length of stroke. Thus: 



Cross-sectional area of the 
bore in square inches 

Volume of cylinder swept 
through by the piston each 
stroke in cubic inches 



ttD* 



xL 



Once the cylinder volume is known, the total 
cylinder volume swept through by the piston of 
a single cylinder compressor each minute in 
cubic inches can be determined by multiplying 
the cylinder volume by the rpm (N). When the 
compressor has more than one cylinder, the 
cylinder volume must also be multiplied by the 
number of cylinders («). In either case, dividing 
the result by 1728 will give the piston displace- 
ment in cubic feet per minute. 



Example 12-1. Calculate the piston dis- 
placement of a two cylinder compressor 
rotating at 1450 rpm, if the diameter of the 
cylinder is 2.S in. and the length of stroke is 2 
in. 

Solution. Substituting in Equation 12-1, 
3.1416 x (2.5)* x 2 x 1455 x 2 
4 x 1728 

= 16.52 cu ft/min 

12-4. Theoretical Refrigerating Capacity. 

The refrigerating capacity of any compressor 
depends upon the operating conditions' of the 
system and, like system capacity, is determined 
by the weight of refrigerant circulated per unit 
of time and by the refrigerating effect of each 
pound circulated.* 

The weight of refrigerant circulated per 
minute by the compressor is equal to the weight 
of the suction vapor that the compressor 
compresses per minute. If it is assumed that the 
compressor is 100% efficient and that the 
cylinder of the compressor fills completely with 
suction vapor at each downstroke of the piston, 
the volume of suction vapor drawn into the 
compressor cylinder and compressed per minute 
will be exactly equal to the piston displacement 
of the compressor. The weight of this volume 
of vapor, which is the weight of refrigerant 
circulated per minute, can be calculated by 
multiplying the piston displacement of the 
compressor by the density of the suction vapor 
at the compressor inlet. 

Once the weight of refrigerant compressed 
per minute by the compressor has been deter- 
mined, the theoretical refrigerating capacity of 
the compressor in tons can be found by multi- 
plying the weight of refrigerant compressed per 
minute by the refrigerating effect per pound and 
then dividing by 200. 

Example 12-2. The compressor in Example 
12-1 is operating on a R-12 system at a suction 
temperature of 20° F. If the suction vapor 
reaching the compressor inlet is saturated and 
if the temperature of the liquid at the refrigerant 
control is 100° F, determine 

(a) the total weight of refrigerant circulated 
per minute 

(6) the theoretical refrigerating capacity of 
the compressor in tons. 

* Since it is the compressor which circulates the 
refrigerant through the system, compressor capacity 
and system capacity are one and the same. 



206 PRINCIPLES OF REFRIGERATION 



Solution 

(a) From Example 12-1, 
piston displacement 
From Table 16-3, den- 
sity of R- 12 saturated 
vapor at 20° F 

Weight of refrigerant 
circulated per minute 

(b) From Table 16-3, 
enthalpy of R-12 satur- 
ated vapor at 20° F 
Enthalpy of R-12 satur- 
ated liquid at 100° F 

Refrigerating effect 
Theoretical refrigerating 
capacity of com- 
pressor 
Theoretical refriger- 
ating capacity in 
tons 



= 16.52 cu ft/min 



= 0.8921 lb/cu ft 
= 16.52 x 0.8921 
= 14.74 lb/min 



= 80.49 Btu/lb 

= 31.16 Btu/lb 
= 49.33 Btu/lb 

= 14.74 x 49.33 
= 727.12 Btu/min 
727.12 



200 
= 3.63 tons 



Since specific volume is the reciprocal of 
density, an alternate method of determining 
the weight of refrigerant circulated per minute 
by the compressor is to divide the piston 
displacement of the compressor by the specific 
volume of the suction vapor at the compressor 
inlet. 

When the volume of vapor to be circulated 
per minute per ton for any given operating 
conditions is known, the capacity of the com- 
pressor in tons for the operating conditions in 
question may be found by dividing the piston 
displacement of the compressor by the volume 
of vapor to be compressed per minute per ton. 

Example 12-3. For the conditions of 
Example 12-2, find (a) the weight of refrigerant 
circulated per minute per ton ; (6) the volume of 
vapor to be compressed per minute per ton; 
and (c) the theoretical refrigerating capacity of 
the compressor in tons. 

Solution 

(a) From Example 12-2, 

refrigerating effect = 49.33 Btu/lb 

Weight of refrigerant _ 200 

circulated per minute ~~ 4933 
Per ton = 4.05 lb/min 

(b) From Table 16-3, 
specific volume of 
R-12 saturated vapor 

at20°F = 1.121 cu ft/lb 

Volume of vapor to be 

compressed per =4.05 x 1.121 

minute per ton = 4.55 cu ft/min 



(c) Piston displacement of 
compressor 

Theoretical refriger- 
ating capacity of 
compressor in tons 



= 16.52 cu ft/min 
_ 16.52 
~ 4.55 
= 3.63 tons 



12-5. Actual Refrigerating Capacity. The 

actual refrigerating capacity of a compressor is 
always less than its theoretical capacity as 
calculated in the previous examples. In the 
preceding examples it has been assumed: (1) 
that at each downstroke of the piston the cylinder 
of the compressor fills completely with suction 
vapor from the suction line and (2) that the 
density of the vapor filling the cylinder is the 
same as that in the suction line. 

If these assumptions were correct, the actual 
refrigerating capacity would be exactly equal to 
the theoretical capacity. Unfortunately, this is 
not the case. Because of the compressibility 
of the refrigerant vapor and the mechanical 
clearance between the piston and the valve plate 
of the compressor, the volume of suction vapor 
filling the cylinder during the suction stroke is 
always less than the cylinder volume swept 
through by the piston. Too, it will be shown 
later that the density of the vapor filling the 
cylinder is less than the density of the vapor 
in the suction line. For these reasons, the actual 
volume of suction vapor at suction line condi- 
tions which is drawn into the cylinder of the 
compressor is always less than the piston 
displacement of the compressor and, therefore, 
the actual refrigerating capacity of the com- 
pressor is always less than its theoretical 
capacity. 

12-6. Total Volumetric Efficiency. The 
actual volume of suction vapor compressed per 
minute is the actual displacement of the com- 
pressor. The ratio of the actual displacement of 
the compressor to its piston displacement is 
known as the total or real volumetric efficiency 
of the compressor. Thus: 



E -* 



x 100 



(12-2) 



where E v = the total volumetric efficiency 

V a = actual volume of suction vapor 

compressed per minute 
V v — the piston displacement of (he 

compressor 



PERFORMANCE OF RECIPROCATING COMPRESSORS 207 



or 



E v = 



Actual weight of suction vapor 

compressed x 100 

Theoretical weight of suction vapor 

compressed 

When the volumetric efficiency of the com- 
pressor is known, the actual displacement and 
refrigerating capacity can be found as follows: 

Ev (12-3) 



V n - K. x 



100 



and 



Actual 



Theoretical 



refrigerating = refrigerating x — ^- (12-4) 
capacity capacity 

Example 12-4. If the volumetric efficiency 
of the compressor in Example 12-3 is 76%, 
determine: (a) the actual volumetric displace- 
ment (b) the actual refrigerating capacity. 

Solution 
(«) From Example 12-1, 

piston displacement 

Actual volumetric 
displacement 
(b) From Example 12-3, 

theoretical refrigerating 

capacity 

Actual refrigerating 
capacity 



16.52 cu ft/min 
16.52 x 0.76 
12.66 cu ft/min 



3.63 tons 
3.63 x 0.76 
2.76 tons 



The actual refrigerating capacity of the 
compressor ' may also be determined as in 
Examples 12-2 and 12-3, if actual displacement 
is substituted for piston displacement. 
12-7. Factors Influencing Total Volumetric 
Efficiency. The factors which tend to limit the 
volume of suction vapor compressed per work- 
ing stroke, thereby determining the volumetric 
efficiency of the compressor, are the following: 

1. Compressor clearance 

2. Wiredrawing 

3. Cylinder heating 

4. Valve and piston leakage 

12-8. The Effect of Clearance on Volumetric 
Efficiency. Because of compressor clearance 
and the compressibility of the refrigerant vapor, 
the volume of suction vapor flowing into the 
cylinder is less than the volume swept through 
by the piston. As previously shown, at the end 
of each compression stroke a certain amount of 
vapor remains in the cylinder in the clearance 



space after the discharge valves close. The 
vapor left in the clearance space has been 
compressed to the discharge pressure and, at 
the beginning of the suction stroke, this vapor 
must be re-expanded to the suction pressure 
before the suction valves can open and allow 
vapor from the suction line to flow into the 
cylinder. The piston will have completed a part 
of its suction stroke and the cylinder will already 
be partially filled with the re-expanded clearance 
vapor before the suction valves can open and 
admit suction vapor to the cylinder. Hence, 
suction vapor from the suction line will fill only 
that part of the cylinder volume which is not 
already filled with the re-expanded clearance 
vapor. 

In Fig. 12-3, V c is the total volume of the 
cylinder when the piston is at the bottom of its 
stroke. V a , which represents the clearance 
volume, is the volume occupied by the clearance 
vapor at the end of the compression stroke. 
The difference between V c and V a then is the 
volume of the cylinder swept through by the 
piston each stroke. On the down stroke of the 
piston, the clearance vapor expands from V a to 
V b before the suction valves open. Therefore, 
the part of the cylinder volume which is filled 
with suction vapor during the balance of the 
suction stroke is the difference between Vj, 
and V e . 

12-9. Theoretical Volumetric Efficiency. 
The volumetric efficiency of a compressor due 
to the clearance factor alone is known as the 
theoretical volumetric efficiency. It can be 
shown mathematically that the theoretical 
volumetric efficiency varies with the amount of 
clearance and with the suction and discharge 
pressures. The reason for this is easily explained, 
12-10. Effect of Increasing the Clearance. 
If the clearance volume of the compressor is 
increased in respect to the piston displacement, 
the percentage of high-pressure vapor remaining 
in the cylinder at the end of the compression 
stroke will be increased. When re-expansion 
takes place during the suction stroke, a greater 
percentage of the total cylinder volume will be 
filled with the re-expanded clearance vapor and 
the volume of suction vapor taken in per stroke 
will be less than when the clearance volume is 
smaller. To obtain maximum volumetric 
efficiency, the clearance volume of a vapor 
compressor should be kept as small as possible. 



208 PRINCIPLES OF REFRIGERATION 



It should be noted that this does not hold 
true for a reciprocating liquid pump. Since a 
liquid is not compressible, the liquid left in the 
clearance space at the end of the discharge 
stroke has the same specific volume as the liquid 
at the suction inlet. Therefore, there is no 
re-expansion of the liquid in the clearance during 
the suction stroke and the volume of liquid 
taken in each stroke is always equal to the 
volume swept by the piston, regardless of 
clearance. 

12-1 1. Variation with Suction and Discharge 
Pressures. Increasing the discharge pressure or 
lowering the suction pressure will have the same 
effect on volumetric efficiency as increasing the 
clearance. If the discharge pressure is increased, 
the vapor in the clearance will be compressed 
to a higher pressure and a greater amount of 
re-expansion will be required to expand it to 
the suction pressure. Likewise, if the suction 
pressure is lowered, the clearance vapor must 
experience a greater re-expansion in expanding 
to the lower pressure before the suction valves 
will open. 

On the other hand, for a constant discharge 
pressure, the amount of re-expansion that the 
clearance vapor experiences before the suction 
valves open diminishes as the suction pressure 
rises. It is evident, then, that the volumetric 
efficiency of the compressor increases as the 
suction pressure increases and decreases as the 
discharge pressure increases. 
12-12. Compression Ratio. The ratio of the 
absolute suction pressure to the absolute 
discharge pressure is called the compression 
ratio. Thus, 

Absolute discharge pressure 

JK — — r-: : . (12-3) 

Absolute suction pressure 
where R = the compression ratio. 

Example 12-5. Calculate the compression 
ratio of a R-12 compressor when the suction 
temperature is 20° F and the condensing tem- 
perature is 100° F. 

Solution. From Table 16-3, 
absolute pressure of R-12 satur- 
ated vapor at 20° F = 35.75 psi 

Absolute pressure of R-12 
saturated vapor at 100° F = 131.6 psi 

Compression ratio _ 131.6 

- 35?75 
= 3.69 



Examination of Equation 12-3 indicates that 
the compression ratio is increased by either 
increasing the discharge pressure or lowering 
the suction pressure, or both. 

In the preceding section it was shown that 
increasing the discharge pressure or lowering 
the suction pressure decreases the volumetric 
efficiency. It follows, then, that when the suction 
and discharge pressures are varied in such a 
direction that the compression ratio is increased, 
the volumetric efficiency of the compressor 
decreases. Likewise, decreasing the compression 
ratio will increase the volumetric efficiency. For 
a compressor of any given clearance, the volu- 
metric efficiency varies inversely with the 
compression ratio. 

12-13. The Effects of Wiredrawing. Wire- 
drawing is defined as a "restriction of area for a 
flowing fluid, causing a loss in pressure by 
(internal and external) friction without the loss 
of heat or performance of work; throttling."* 

In order to have a flow of vapor from the 
suction line through the suction valves into the 
compressor cylinder, there must be a pressure 
differential across the valves sufficient to over- 
come the spring tension of the valves and valve 
weight and inertia. This means that the 
suction vapor experiences a mild, throttling 
expansion or drop in pressure as it flows 
through the suction valves and passages of the 
compressor. Therefore, the pressure of the 
suction vapor filling the cylinder of the com- 
pressor is always less than the pressure of the 
vapor in the suction line. As a result of the 
expanded condition of the vapor filling the 
cylinder, the volume of suction vapor taken in 
from the suction line each stroke is less than 
if the vapor filling the cylinder was at the 
suction line pressure. 

A similar pressure differential is required 
across the discharge valves in order to cause the 
discharge vapor to flow through the valves 
into the condenser. To provide the necessary 
pressure differential across the discharge valves, 
the vapor in the cylinder must be compressed to 
a pressure somewhat higher than die actual 
condensing pressure. The vapor left in the 
clearance space at the end of the di charge 
stroke will be at this higher pressure. To re- 
expand from this higher pressure during the 

* Asre Data Book, 1957-58 (page 39-27). 



PERFORMANCE OF RECIPROCATING COMPRESSORS 209 



suction stroke, the clearance vapor must suffer 
a greater amount of re-expansion than if it had 
been compressed only to the condensing pres- 
sure. As a result of the greater expansion of 
the clearance vapor, a larger portion of the 
cylinder volume is filled with the re-expanded 
clearance vapor during the down stroke of the 
piston and the amount of suction vapor drawn 
in from the suction line is reduced. 

Unlike the other factors which determine 
volumetric efficiency, wiredrawing is not directly 
affected by the compression ratio. In general, 
wiredrawing is a function of the velocity of the 
refrigerant vapor flowing through the valves and 
passages of the compressor. As the velocity of 
the vapor through the valves is increased, the 
effect of wiredrawing increases. 

The refrigerant velocity through the valves of 
a compressor depends upon the design of the 
valves, the refrigerant used, and the speed of the 
compressor. 

Wiredrawing is greatest for those refrigerants 
having the greatest specific volumes and the 
lowest latent heat values because the volume 
of vapor circulated per ton of refrigerating 
capacity is greater. This accounts for the large 
wiredrawing effect associated with R-12. 

Increasing the speed of the compressor 
increases the piston displacement. Hence, the 
velocity of the vapor through the valves and the 
effects of wiredrawing are increased as the rpm 
are increased. 

12-14. The Effects of Cylinder Heating. 
Another factor which tends to reduce the 
volumetric efficiency of the compressor is the 
heating of the suction vapor in the compressor 
cylinder. The suction vapor entering the 
compressor cylinder is heated by heat con- 
ducted from the hot cylinder walls and by 
friction which results from the turbulence of the 
vapor in the cylinder and from the fact that the 
refrigerant vapor is not a perfect gas. The 
heating causes the vapor to expand after 
entering the cylinder so mat a smaller weight of 
vapor will fill the cylinder and thereby still 
further reduce the volume of vapor taken in 
from the suction line. 

Cylinder heating increases as the compression 
ratio increases. At high compression ratios, 
the work of compression is greater and the 
discharge temperature is higher. This causes 
a rise in the temperature of the cylinder walls 



and other compressor parts so that the transfer 
of heat to the suction vapor occurs at a higher 
rate. 

12-15. The Effect of Piston and Valve Leak- 
age. Any back leakage of gas through either the 
suction or discharge valves or around the piston 
will decrease the volume of vapor pumped by 
the compressor. Because of precision manu- 
facturing processes, there is very little leakage of 
gas around the pistons of a compressor in good 
condition. However, since it is not possible to 
design valves that will close instantaneously, 
there is always a certain amount of back 
leakage of gas through the suction and discharge 
valves. 

As the pressure in the cylinder is lowered at 
the beginning of the suction stroke, a small 
amount of high-pressure vapor in the head of 
the compressor will leak back into the cylinder 
before the discharge valves can close tightly. 
Similarly, at the start of the compression stroke, 
some of the vapor in the cylinder will flow 
back through the suction valves into the, suction 
line before the suction valves can close. 

To assure prompt closing of the valves, both 
the suction and discharge valves are usually 
constructed of lightweight materials and are 
slightly spring loaded. However, since the spring 
tension increases wiredrawing, the amount of 
spring loading is critical. 

For any given compressor, the amount of 
backleakage through the valves is a function of 
the compression ratio and the speed of the 
compressor. The higher the compression ratio, 
the greater is the amount of valve leakage. 
The effect of compressor speed on valve 
leakage is discussed later. 
12-16. Determining the Total Volumetric 
Efficiency. The combined effects of all of the 
foregoing factors on the volumetric efficiency of 
the compressor varies with the design of the 
compressor and with the refrigerant used. 
Furthermore, for any one compressor the 
volumetric efficiency is not a constant amount; 
it changes with the operating conditions of the 
system. Therefore, the total volumetric effi- 
ciency of a compressor is difficult to predict 
mathematically and can be determined with 
accuracy only by actual testing of the compressor 
in a laboratory. 

However, the results of such tests indicate 
that the volumetric efficiency of any one 



210 PRINCIPLES OF REFRIGERATION 



Volumetric efficiency 



















































































































































































































2 3 4 5 6 7 8 9 10 11 12 13 14 
Compression ratio 

Fig. 12-4. Effect of compression ratio on volumetric 
efficiency of R-12 compressor. 



compressor is primarily a function of the 
compression ratio and, for any given com- 
pression ratio, remains practically constant, 
regardless of the operating range. It has been 
determined also that compressors having the 
same design characteristics will have approxi- 
mately the same volumetric efficiencies, regard- 
less of the size of the compressor. 

The relationship between the compression 
ratio and the volumetric efficiency of a typical 
R-12 compressor is illustrated by the curve in 
Fig. 12-4. In addition, in order to facilitate 
future calculations, the average volumetric 
efficiencies of a group of typical R-12 com- 
pressors at various compression ratios are 
given in Table 12-1. The values given are for 
compressors ranging in size from 5 to 25 hp. 
Smaller compressors will have slightly lower 
efficiencies, whereas larger compressors will 
have slightly higher efficiencies. 
12-17. Variation in Compressor Capacity 
with Suction Temperature. Compressor per- 
formance and cycle efficiency will vary consider- 
ably with the operating conditions of the system. 
The most important factor governing the 
capacity of the compressor is the vaporizing 
temperature of the liquid in the evaporator, 
that is, the suction temperature. The large 
variations in compressor capacity which accom- 
pany changes in the operating suction tempera- 
ture are primarily a result of a difference in the 
density of the suction vapor entering the 
suction inlet of the compressor. The higher the 
vaporizing temperature of the liquid in the 
evaporator, the higher is the vaporizing pressure 
and the greater is the density of the suction 
vapor. Because of the difference in the density 



of the suction vapor, each cubic foot of vapor 
compressed by the compressor will represent a 
greater weight of refrigerant when the suction 
temperature is high than when the suction 
temperature is low. This means that for any 
given position displacement, the weight of 
refrigerant circulated by the compressor per 
unit of time increases as the suction temperature 
increases. 

The effect of suction temperature on com- 
pressor capacity is best illustrated by an actual 
example. 

Example 12-6. Assuming 100% efficiency, 
if the liquid reaches the refrigerant control at 
100° F in each case, determine the weight 
of refrigerant circulated per minute and the 
theoretical refrigerating capacity of the com- 
pressor in Example 12-1 when operating at each 
of the following suction temperatures: (a) 10° F 
and (b) 40° F. 

Solution 
(a) From Table 16-3, 
density of R-12 satur- 



ated vapor at 10° F 


0.7402 lb/cu ft 


From Example 12-1, 
piston displace- 
ment 


= 16.52 cu ft/min 


Weight of refrigerant 
circulated per 
minute at 10° F 
suction 


= 16.52 x 0.7402 
= 12.23 lb/min 


From Table 16-3, 
enthalpy of R-12 
saturated vapor at 
10° F 


= 79.36 Btu/lb 


Enthalpy of R-12 
liquid at 100° F 


= 31.16 Btu/lb 


Refrigerating effect 


= 48.20 Btu/lb 


Theoretical refriger- 
ating capacity of 
compressor at 10° F 
suction, Btu/min 


= 12.23 x 48.20 
= 589.49 Btu/min 


Theoretical refrigerat- 
ing capacity in tons 

(b) From Table 16-3, 
density of R-12 satur- 
ated vapor at 40° F 


589.49 

200 

= 2.95 tons 

= 1.263 lb/cu ft 


From Example 12-1, 
piston displace- 
ment 


= 16.52 cu ft/min 



PERFORMANCE OF RECIPROCATING COMPRESSORS 211 



Weight of refrigerant 
circulated per minute 


= 16.52 x 1.263 


at 40° F suction 


= 20.86 lb/min 


From Table 16-3, 




enthalpy of R-12 
saturated vapor at 
40° F 


= 82.71 Btu/lb 


Enthalpy of R-12 
liquid at 100° F 


= 31.16 Btu/lb 


Refrigerating effect 


= 51.55 Btu/lb 


Theoretical refriger- 




ating capacity of 
compressor at 40° F 
suction, Btu/min 


= 20.86 x 51.55 
= 1075.33 Btu/min 


Theoretical refriger- 


1075.33 


ating capacity in 


200 


tons 


= 5.38 tons 



In analyzing the results of Example 12-6 the 
following observations are of interest: 

1. Although the piston displacement of the 
compressor is the same in each case, the weight 
of refrigerant circulated per minute by the 
compressor increases from 12.23 lb/min to 
20.86 lb/min when the operating suction 
temperature is raised from 10° F to 40° F. The 
increase in the weight of refrigerant circulated 
results entirely from the greater density of the 
suction vapor entering the suction inlet of the 
compressor. In this instance, the percentage 
increase in the weight of refrigerant circulated is 



20.86 - 12.23 



12.23 



x 100 =70.5% 



2. The theoretical refrigerating capacity of 
the compressor at the 10° F suction temperature 
is 2.95 tons, whereas at the 40° F suction 
temperature, the capacity increases to 5.38 tons. 
This represents an increase in refrigerating 
capacity of 

5.38 - 2.95 
2^5 x 100 = 82.3% 

Although the increased density of the suction 
vapor at the higher suction temperature 
accounts for the greater part of the increase in 
compressor capacity, it is not the only reason 
for it. As indicated, the increase in the weight 
of refrigerant circulated is only 70.5%, whereas 
the total increase in compressor capacity is 
82.3%. The additional 1 1.8 %gain in capacity is 
brought about by an increase in the refrigerating 



effect of each pound of refrigerant circulated. 
Although the actual gain in refrigerating effect 
per pound is only 6.95 %, when this increase is 
applied to the entire weight of refrigerant 
circulated at the higher suction temperature, the 
net gain in capacity over the original capacity 
which can be attributed to the greater refriger- 
ating effect is 11.8% (1.705 x 0.0695 = 1.823 
and 1.283 - 1.705 = 0.118 or 11.8%). 

The actual variation in compressor capacity 
with changes in suction temperature is more 
pronounced than that indicated by theoretical 
computations. That is, the change in the actual 
compressor capacity with variations in suction 
temperature is always greater than the change 
in the theoretical capacity. The reason for this 
is that the compression ratio changes as the 
suction temperature changes. When the vapor- 
izing temperature increases while the condensing 
temperature remains constant, the compression 
ratio is decreased and the volumetric efficiency 
of the compressor is improved. Hence, at the 
higher suction temperature, in addition to 
pumping a greater weight of refrigerant per unit 
of volume, the volume of vapor pumped by the 
compressor is also larger because of the im- 
proved efficiency. 

Example 12-7. Assuming that the satur- 
ated discharge temperature is 100° F, determine 
the actual refrigerating capacity of the com- 
pressor in Example 12-6 when operating at each 
of the suction temperatures in question. 



Solution 




(a) From Table 16-3, 




absolute pressure corre- 




sponding to 100° F 




saturation temperature 


131.6 psi 


Absolute pressure corre- 




sponding to 10° F 




saturation temperature 


= 29.35 psi 


Compression ratio 


131.6 




29.35 




= 4.47 


From Table 12-1, 




volumetric efficiency 


= 76.3% 


From Example 12-6, 




theoretical refrigerating 




capacity at 10° F 




suction 


= 2.95 tons 


Actual refrigerating 




capacity at 10° F 


= 2.95 x 0.763 


suction 


= 2.22 tons 



212 PRINCIPLES OF REFRIGERATION 



(b) From Table 16-3, 
absolute pressure corre- 
sponding to 100° F satu- 
ration temperature 
Absolute pressure 
corresponding to 40° F 



= 131.6 psi 



saturation temperature 


= 51.68 psi 


Compression ratio 


131.6 




31.68 




= 2.55 


From Table 12-1, 




volumetric efficiency 


= 85.7% 


From Example 12-6, 
theoretical refrigerating 
capacity at 40° F 
suction 


= 5.38 tons 


Actual refrigerating 
capacity at 40° F 
suction 


= 5.38 x 0.857 
= 4.61 tons 



Whereas the theoretical increase in com- 
pressor capacity is only 82.3%, the actual 
increase in refrigerating capacity is 



4.61 -2.22 



2.22 
= 107.7% 



x 100 



12-18. Effect of Condensing Temperature 
on Compressor Capacity. In general, the 
refrigerating capacity of the compressor de- 
creases as the condensing temperature increases 
and increases as the condensing temperature 
decreases. The effect that the condensing 
temperature has on compressor efficiency and 
capacity can be evaluated by comparing the 
results of the following example with those of 
Examples 12-6 and 12-7. 

Example 12-8. Determine the theoretical 
and actual refrigerating capacities of the com- 
pressor in Example 12-1 for each of the two 
vaporizing temperatures given in Examples 
12-6 and 12-7, if the condensing temperature in 
each case is 120° F rather than 100° F. 

Solution 
(a) For the 10° F vaporizing temperature. 
From Example 12-1, 
piston displacement 
of compressor 

From Table 16-3, den- 
sity of R-12 saturated 
vapor at 10° F = 0.7402 lb/cu ft 



= 16.52 cu ft/min 



Theoretical weight of 
refrigerant circulated 
per minute by corn- 
compressor 

Refrigerating effect per 
pound at 10° F 
vaporizing and 120° F 
condensing 

Theoretical refriger- 
ating capacity of 
compressor 

Theoretical refriger- 



= 16.52 x 0.7402 
= 12.23 lb 



= 43.20 Btu/lb 

= 12.23 x 43.20 
= 527.34 Btu/min 

527.34 



ating capacity in tons 


200 




= 2.64 tons 


From Table 16-3, 




absolute suction 




pressure 


= 29.35 psi 


Absolute discharge 




pressure 


= 171.8 psi 


Compression ratio 


171.8 



29.35 
= 5.85 



= 66.5% 

= 2.645 x 
= 1.76 



From Table 12-1, 
volumetric efficiency 

Actual refrigerating 
capacity in tons 

(ft) For the 40° F suction temperature. 
From Example 12-1, 
piston displacement 
of compressor 

From Table 16-3, den- 
sity of R-12 saturated 
vapor at 40° F 

Theoretical weight of 
refrigerant circulated 
per minute by 
compressor 



0.665 



= 16.52 cu ft/min 



= 1.263 lb/cu ft 



= 16.52 x 1.263 
= 20.86 lb 



Refrigerating effect per 
pound at 40° F 
evaporating and 
condensing 120° F 

Theoretical refriger- 
ating capacity of 
compressor 

Theoretical refriger- 
ating capacity in tons 

From Table 16-3, 
absolute suction 
pressure 

Absolute discharge 
pressure 



= 46.55 Btu/lb 

= 20.86 x 46.55 
= 971 Btu/min 

971 
~200 
= 4.85 



= 51.68 psi 
= 171.8 psi 



PERFORMANCE OF RECIPROCATING COMPRESSORS 213 



Compression ratio 



From Table 12-1, 
volumetric efficiency 

Actual refrigerating 
capacity in tons 



171.8 
51.68 
3.32 



78.5% 

4.85 x 0.785 
3.81 



Examining first the 10° F cycle, notice that 
raising the condensing temperature from 100° F 
to 120° F reduces the theoretical refrigerating 
capacity of the compressor from 2.95 tons to 
2.64 tons and the actual capacity from 2.22 tons 
to 1.76 tons. 

Since a 100% efficient compressor is assumed 
to displace a theoretical volume of vapor equal 
to its piston displacement and since the density 
of the suction vapor entering the compressor 
for any one vaporizing temperature is always 
the same regardless of the condensing tempera- 
ture, the theoretical weight of refrigerant 
displaced by the compressor is the same at all 
condensing temperatures, and therefore the 
theoretical refrigerating capacity of the com- 
pressor for any condensing temperature depends 
only upon the refrigerating effect per pound of 
refrigerant circulated. Hence, the difference 
in the theoretical refrigerating capacity of the 
compressor at the two condensing temperatures 
results entirely from the difference in the 
refrigerating effect per pound. 

The reduction in actual compressor capacity 
may be attributed to several factors: (1) a 
reduction in the refrigerating effect per pound 
and (2) a reduction in the volumetric efficiency 
of the compressor. 

Increasing the condensing temperature while 
the suction temperature remains constant 
increases the compression ratio and reduces 
the volumetric efficiency of the compressor so 
that the actual volume of vapor displaced by the 
compressor per unit of time decreases. There- 
fore, even though the density of the vapor 
entering the compressor remains the same at 
all condensing temperatures, the actual weight 
of refrigerant circulated by the compressor per 
unit of time decreases because of the reduction 
in the quantity of vapor handled. 

Increasing the condensing temperature in- 
creases the isentropic discharge temperature. In 
this instance, it is interesting to note (Fig. 7-7) 
that the increase in the isentropic discharge 



temperature is somewhat greater than that in the 
condensing temperature. Whereas the increase 
in condensing temperature is only 20° F 
(120° - 100°), the increase in the discharge 
temperature is 23.5° F (137.5° - 114°). This is 
accounted for by the greater work of com- 
pression at the higher compression ratio. Had 
the condensing temperature been increased in 
such a way that the compression ratio does not 
change (by increasing the suction temperature 
in proportion), the increase in the discharge 
temperature would have been approximately 
the same as that in the condensing tempera- 
ture. 

High discharge temperatures are undesirable 
and are to be avoided whenever possible. The 
higher the discharge temperature, the higher is 
the average temperature of the cylinder walls 
and the greater is the superheating of the 
suction vapor in the compressor cylinder. In 
addition to its adverse effect of compressor 
efficiency, high discharge temperatures tend to 
increase the rate of acid formation in the 
system, cause carbonization of the oil in the 
head of the compressor, and produce other 
effects detrimental to the equipment. 

The loss of compressor efficiency and capacity 
resulting from an increase in the condensing 
temperature of the cycle is more serious when 
the suction temperature of the cycle is low than 
when the suction temperature is high. The 
desirability of operating a refrigerating system 
at the lowest practical condensing temperature 
has already been pointed out. This is of par- 
ticular importance when the suction temperature 
of the cycle is low and the compressor is already 
operating at a relatively low efficiency. 

When the cycle is operating at a 40° F 
vaporizing temperature, increasing the con- 
densing temperature from 100° F to 120° F 
reduces the theoretical capacity of the com- 
pressor from 5.38 tons to 4.85 tons and the 
actual compressor capacity from 4.61 tons to 
3.81 tons. The loss in theoretical capacity is 



5.38 - 4.85 



5.38 



x 100 = 10% 



The loss in actual compressor capacity amounts 
to 

4.61 - 3.81 

4 . 61 x 100 = 17.4% 



214 PRINCIPLES OF REFRIGERATION 



For the 10° F cycle, the loss in theoretical 
compressor capacity is 

2.95 - 2.64 



2.95 



x 100 = 10.5% 



and the loss in actual compressor capacity is 
2.55 - 1.76 



2.55 



x 100 = 31% 



Note that the loss in theoretical capacity 
brought about by increasing the condensing 
temperature is approximately the same for 
both suction temperatures, whereas the loss in 
actual compressor capacity is much greater at 
the lower suction temperature. To a great 
extent, it is the loss in volumetric efficiency that 
causes the marked decrease in the actual 
capacity of the compressor at the higher 
condensing temperature. The change in volu- 
metric efficiency for a given change in condens- 
ing temperature becomes greater as the suction 
temperature of the cycle decreases. This 
accounts for the greater effect that a change in 
condensing temperature has on compressor 
capacity when the suction temperature is low. 
12-19. Compressor Horsepower. The theore- 
tical horsepower required to drive the compres- 
sor may be found by multiplying the actual 
refrigerating capacity of the compressor in tons 
by the theoretical horsepower required per ton 
for the operating conditions in question. 

Example 12-9. Find the theoretical horse- 
power required to drive the compressor in 
Example 12-4. 

Solution. From Example 
12-4, actual refrigerating 
capacity in tons = 2.76 tons 

From Fig. 7-9, theoretical 
horsepower required per ton = 0.965 hp 

The theoretical horsepower = 2.76 x 0.965 
required to drive the com- 
pressor = 2.66 hp 

Notice that it is the actual refrigerating 
capacity of the compressor, rather than the 
theoretical refrigerating capacity, which must 
be used in determining the theoretical horse- 
power requirements of the compressor. 

The theoretical horsepower as calculated in 
the preceding example is only an indication of 
the power which would be required by a 100% 



efficient compressor operating on an ideal 
compression cycle and does not represent the 
actual total horsepower which must be delivered 
to the shaft of the compressor. In actual 
practice, there are certain losses in power which 
accrue because of the mechanical friction in the 
compressor and because of the deviation of an 
actual compression cycle from the ideal com- 
pression cycle. Naturally, additional power 
must be supplied to the compressor to offset 
these losses. Therefore, the actual power 
required by a compressor will always be greater 
than the theoretical computations indicate. 
12-20. Variation in Compressor Horse- 
power with Suction Temperature. Although 
the horsepower per ton of refrigerating capacity 
diminishes as the suction temperature rises, the 
horsepower required by the compressor may 
either increase or decrease, depending upon 
whether the work done by the compressor 
increases or decreases. 

The total amount of work done by the 
compressor per unit of time in compressing the 
vapor and, hence, the power required to drive 
the compressor, is the function of only two 
factors : (1) the work of compression per pound 
of vapor compressed and (2) the weight of vapor 
compressed per unit of time. 

The amount of work which is done in 
compressing the vapor from the suction 
pressure to the discharge pressure varies with 
the compression ratio. The greater the com- 
pression ratio, the greater is the work of 
compression. Therefore, when the suction 
temperature is raised while the condensing 
temperature remains the same, the compression 
ratio becomes smaller and the work of com- 
pression per pound is reduced. However, at the 
same time, because of the greater density of the 
suction vapor, the weight of vapor compressed 
by the compressor per unit of time increases. 
Since the saving in work done resulting from 
the reduction in the work per pound is seldom 
sufficient to outweigh the increase in the work 
of the compressor because of the increase in the 
weight of vapor compressed, raising the suction 
temperature will usually increase the power 
requirements of the compressor. 

Example 12-10. Compute the theoretical 
horsepower required by the compressor in 
Example 12-7 at each of the suction tempera- 
tures listed. 



PERFORMANCE OF RECIPROCATING COMPRESSORS 215 



Solution 
(a) From Example 12-7, 
actual refrigerating 
capacity in tons at 10° F 
suction temperature 
From Fig. 7-9, theoreti- 
cal horsepower per ton 
at 10° suction and 
100° F condensing 
Theoretical horsepower 
of compressor at 10° F 
suction 
(6) From Example 12-7, 
actual refrigerating 
capacity in tons at 40° F 
suction temperature 
From Fig. 7-9, theoreti- 
cal horsepower per ton 
at 40° F suction and 
100° F condensing 
Theoretical horsepower 
of compressor at 40° F 
suction 



= 2.22 tons 



= 1.13 hp 

= 2.22 x 1.13 
= 2.51 hp 



= 4.61 tons 



= 0.683 hp 



= 4.61 x 0.683 
= 3.15 hp 

Although the horsepower per ton decreases 
39.5 % as the suction temperature is raised from 
10° F to 40° F, because of the increase in the 
refrigerating capacity of the compressor, the 
horsepower required by the compressor in- 
creases from 2.51 hp to 3.15 hp. This represents 
an increase in the power required of 

3.15 -2.51 

— 2 1T -X100=21% 

The increase in compressor horsepower with 
the suction temperature is relatively small in 

7 



Fig. 12-5. Curves illustrate the 
effects of suction temperature 
on the capacity and horse- 
power of reciprocating com- 
pressors. 



comparison to the increase in compressor 
capacity. In this instance, for a 30° F rise in 
suction temperature, the capacity of the 
compressor increased 107%, whereas com- 
pressor horsepower increased only 21 %. The 
average increase in compressor capacity per 
degree of rise in suction temperature is 107%/ 
30° F or 3.21 %, whereas the increase in horse- 
power amounts to only 0.7 % per degree of rise. 

The relationship between compressor capacity 
and the horsepower of the compressor at 
various suction temperatures is shown by the 
curves in Fig. 12-5. The curves are for a typical 
R-12 compressor operating at a constant con- 
densing temperature of 100° F. 

As shown by the curve in Fig. 12-5 the 
horsepower required by a R-12 compressor 
increases as the suction temperature increases 
up to a certain point at which the horsepower 
required by the compressor is at a maximum. 
On reaching this point, if the suction temperature 
is further increased, the horsepower required 
by the compressor diminishes. This is not true, 
however, for compressors using ammonia as a 
refrigerant. For compressors using ammonia, 
the horsepower does not reach a maximum 
value, but continues to increase indefinitely as 
the suction temperature increases. 

The suction temperature at which the horse- 
power required by a R-12 compressors reaches 
a maximum depends upon the condensing 
temperature and increases as the condensing 
temperature increases. 

7 













































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f& 


*&* — 














&sf> 




>? 


&y 














A 


^s 


















^2** 


S^r, 


w fo» 























-40' -30' -20° -10* 0' 10* 20° 30* 40° 50* 
Suction temperature 



216 PRINCIPLES OF REFRIGERATION 



42 
3.0 

-.2.8 

c 

e 

£2.6 
o 

a 

5 2.4 
12 

?.o 












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^ 












^^ 














^s„ 
















s sJS>- 










V**™ 


&**2- 


vx^^^-J 

























80° 



90' 100* 110- 

Condensing temperature 



12-21. The Effect of Condensing Tempera- 
ture on Compressor Horsepower. The 

curves in Fig. 12-6 illustrate the relationship 
between the horsepower required per ton of 
refrigerating capacity, the actual refrigerating 
capacity of the compressor, and the horsepower 
required by the compressor at various condensing 
temperatures when the suction temperature is 
keptconstant. Note that, although the theoretical 
horsepower required per ton increases as the 
condensing temperature increases, the theoretical 
horsepower required by any one compressor will 
not increase in the same proportion. This is true 
because the decrease in the refrigerating capa- 
city of the compressor which is coincident with 
an increase in the condensing temperature will 
offset to some extent the increase in the horse- 
power per ton. 

For instance, according to Fig. 7-9, for a cycle 
operating at a 10° F vaporizing temperature, 
the theoretical horsepower required per ton 
increases from 1 . 1 3 to 1 .52 when the condensing 
temperature of the cycle is increased from 100° F 
to 120° F. At the same time, Example 12-10 
illustrates that the actual refrigerating capacity 
of one particular compressor drops from 2.22 
tons to 1.76 tons when the condensing tempera- 
ture is raised from 100° F to 120° F. The 
theoretical horsepower required by the com- 
pressor at the 100° F condensing temperature is 

1.13 x 2.22 = 2.51 hp 

For the 120° F condensing temperature, the 
theoretical horsepower required by the com- 
pressor is 

1.52 x 1.76 =2.68hp 



4.6 



4.5 



4-4" 



4.3 -S - 



4.2 | 
a. 
E 
o 

4.1 <-> 



4.0 
120* 



2.6 

c 

s 

2.2 I 

i.8 i 

o 
1.4 1 

m 
1.0 



Fig. 12-6. Curves illustrate 
the effects of condensing tem- 
perature on capacity and 
horsepower of reciprocating 
compressors. 



12-22. Brake Horsepower. The total horse- 
power which must be supplied to the shaft of the 
compressor is called the brake horsepower and 
may be computed from the theoretical horse- 
power by application of a factor called over-all 
compressor efficiency. The over-all efficiency is 
an expression of the relationship of the theoreti- 
cal horsepower to the brake horsepower in 
percent. Written as an equation, the relationship 
is 



Thp 

E °=w P xl0 ° 



and 



Bhp = 



Thp 
E o l\00 



(12-6) 



(12-7) 



where E = the over-all efficiency in percent 
Thp = the theoretical horsepower 
Bhp = the brake horsepower 

Example 12-11. Determine the brake 
horsepower required by the compressor in 
Example 12-14, if the over-all efficiency of the 
compressor is 80%. 



Solution. From Example 12-4, 




Thp 


= 3.12 hp 


Applying Equation 12-7, the 


Thp 


Bhp 


E 




3.12 




086 




= 3.9 hp 



The over-all efficiency is sometimes broken 
down into two components : (1) the compression 
efficiency and (2) die mechanical efficiency. In 
such cases, the relationship is 

(12-8) 



PERFORMANCE OF RECIPROCATING COMPRESSORS 217 



where E e = the compression efficiency in per- 
cent 
E m = the mechanical efficiency in percent 



so that 



Bhp = 



Thp 



E c X E„ 



(12-9) 



The compression efficiency of a compressor is 
a measure of the losses resulting from the devia- 
tion of the actual compression cycle from the 
ideal compression cycle, whereas the mechanical 
efficiency of the compressor is a measure of the 
losses resulting from the mechanical friction in 
the compressor. The principal factors which 
bring about the deviation of an actual com- 
pression cycle from the ideal compression cycle 
are: (1) wiredrawing, (2) the exchange of heat 
between the vapor and the cylinder walls, and 
(3) fluid friction due to the turbulence of the 
vapor in the cylinder and to the fact that the 
refrigerant vapor is not an ideal gas. Notice 
that the factors which determine the compression 
efficiency of the compressor are the same as those 
which influence the volumetric efficiency. As a 
matter of fact, for any one compressor, the 
volumetric and compression efficiencies are 
roughly the same and they vary with the com- 
pression ratio in about the same proportions. 
For this reason, the brake horsepower required 
per ton of refrigerating capacity can be approxi- 
mated with reasonable accuracy by dividing the 
theoretical horsepower per ton by the volu- 
metric efficiency of the compressor and then 
adding about 10% to offset the power loss due 
to the mechanical friction in the compressor. 
Written as an equation, 



Bhp = 



M(h d -h e ) x 1.1 
42.42 x E„ 



(12-10) 



Since the relationship between the various 
factors which influence the compression effi- 
ciency are difficult to evaluate mathematically, 
the compression efficiency of a compressor can 
be determined accurately only by actual testing 
of the compressor. 

12-23. Indicated Horsepower. A device fre- 
quently used to determine the compression 
efficiency is the indicator diagram. An indicator 
diagram is a pressure-volume diagram of the 
actual compression cycle of the compressor 
which is produced during the actual testing of 
the compressor. 



A theoretical indicator diagram for an ideal 
compression cycle is shown in Fig. 12-7. It has 
been illustrated previously that the area under 
a process diagram on a pressure-volume chart 
is a measure of the work of the process. In 
Fig. 12-7, notice that the area dDCd represents 
the work done by the piston in compressing the 
vapor during the isentropic process CD, and 
that the area aADba represents the work done 
by the piston in discharging the vapor from the 
cylinder during the constant pressure process 
DA, whereas the area aABa represents the work 
done back on the piston by the vapor during 
the isentropic re-expansion (of the clearance 
vapor) process AB. Since the work of process 
AB is work given back to the piston by the 
fluid, the net work input to the compression 
cycle is the sum of the work of processes CD 
and DA, less the work of process AB. Therefore, 
the net work of the compression cycle is 
represented by the area BADCB, the total area 
enclosed by the cycle diagram. 

The work of the compression cycle as deter- 
mined from the indicator diagram is called the 
indicated work and the horsepower computed 
from the indicated work is called the indicated 
horsepower. 

Since the indicator diagram illustrated in 
Fig. 12-7 is a theoretical indicator diagram of 
an ideal compression cycle, the indicated work 
is the work of an ideal compression cycle and 
the indicated horsepower computed from the 
indicated work would, of course, be exactly 




Volume 



Fig. 12-7. Theoretical indicator diagram for an ideal 
compression cycle. 



218 PRINCIPLES OF REFRIGERATION 



l«J 




D 








_4l^ 






























Q 


1 "' 

1 


1 






1Q 


1 
1 



The relationship of the indicated horsepower 
to the theoretical horsepower is 



V b 



V c ' 
Volume 



Fig. 12-8. Theoretical indicator diagram for an 
actual compression cycle. 



equal to the theoretical horsepower. However, 
in actual practice, since the indicator diagram 
reproduces the true paths of the various 
processes which make up the actual compression 
cycle, the indicated work of the diagram is an 
accurate measure of the actual work of the 
compression cycle and, therefore, the ^indicated 
horsepower computed on the basis of the 
indicated work is the actual horsepower 
required to do the work of the actual com- 
pression cycle. 

Care should be taken not to confuse indi- 
cated horsepower with brake horsepower. 
Although the indicated horsepower includes 
the power required to offset the losses resulting 
from the deviation of an actual compression 
cycle from the ideal cycle, it does not include 
the power required to overcome the losses 
resulting from the mechanical friction in the 
compressor. In other words, the indicated 
horsepower takes into account the compression 
efficiency but not the mechanical efficiency. 
Hence, brake horsepower differs from indicated 
horsepower in that the brake horsepower 
includes the power required to overcome the 
mechanical friction in the compressor, whereas 
the indicated horsepower does not. The 
horsepower necessary to overcome the mechani- 
cal friction in the compressor is sometimes 
referred to as the friction horsepower (Fhp), so 
that 

(12-11) 



Thp 



(12-12) 



Bhp = Ihp + Fhp 



An indicator diagram of an actual com- 
pression cycle is shown in Fig. 12-8. The area 
ABCD, enclosed by the cycle diagram, is, of 
course, a measure of the work of the cycle. 
An ideal cycle, AB'CD', is drawn in for com- 
parison. Pressures P 1 and P 2 represent the 
pressure of the vapor entering and leaving the 
compressor. The areas above line P 2 and 
below line P x represent the increased work of 
the cycle due to wiredrawing. Notice that at 
the end of the suction and discharge strokes 
(points C and A), the piston velocity diminishes 
to zero and the pressure of the vapor tends to 
return to P x and P 2 , respectively. The other 
deviations from the ideal cycle represent the 
losses resulting from the heating of the vapor 
in the compressor cylinder. Line BC indicates 
the approximate volume of the suction vapor at 
the end of the suction stroke, whereas line BC 
represents the approximate volume of this same 
weight of vapor in the suction line. The 
deviation of the actual compression process 
from the isentropic can be seen by comparing 
the actual compression path CD to the isen- 
tropic path CD'. 

The direction of the periodic heat transfer 
between the vapor and the cylinder walls at 
various times and points in the cycle is indi- 
cated by the arrows. The arrows pointing in 
denote heat transfer from the cylinder walls to 
the vapor, whereas arrows pointing out indicate 
heat transfer from the vapor to the cylinder 
walls. 

The temperature of the cylinder walls of the 
compressor will fluctuate around some mean 
value which is between the suction and dis- 
charge temperatures of the vapor. During the 
latter part of the re-expansion process, during 
the period in which the vapor is being admitted 
to the cylinder, and during the initial part of the 
compression stroke, the cylinder wall tempera- 
ture is greater than the vapor temperature and 
heat passes from the cylinder walls to the vapor. 
During the latter part of the compression stroke, 
during the discharge period, and during the 
early part of the suction stroke, the temperature 



PERFORMANCE OF RECIPROCATING COMPRESSORS 219 



of the vapor exceeds the cylinder wall tempera- 
ture and heat passes from the vapor to the 
cylinder walls. 

12-24. Isothermal vs. Isentropic Compres- 
sion. Reference to Fig. 12-9 will show that if 
the compression process in the compressor was 
isothermal rather than isentropic the net work 
of the compression cycle would be reduced even 
though the work of the compression process 
itself is greater for isothermal compression than 
for isentropic compression. The reduction in 
the work of the cycle which would be realized 
through isothermal compression is indicated by 
the crosshatched area in Fig. 12-9. 

Isothermal compression is not practical for a 
refrigeration compressor since it would result 
in the discharge of saturated liquid from the 
compressor. Furthermore, if a cooling medium 
were available at a temperature low enough to 
cool the compressor sufficiently to produce 
isothermal compression, the cooling medium 
could be used directly as the refrigerant and 
there would be no need for the refrigeration 
cycle. 

12-25. Water-Jacketing the Compressor 
Cylinder. Any heat which is given up by the 
compressor cylinder to some external cooling 
medium represents, in effect, heat given up by 
the vapor during the compression process. 
Cooling of the vapor during compression causes 
the path of the compression process to shift 
from the isentropic path toward an isothermal 
path. Of course, the greater the amount of 
cooling, the greater will be the shift toward the 
isothermal. 

If the temperature of the air surrounding the 
compressor were exactly the same as the 
temperature of the compressor cylinder, there 
would be no transfer of heat from the cylinder 
to the air and any heat given up by the vapor 
to the cylinder would be eventually reabsorbed 
by the vapor and the compression process 
would be approximately adiabatic. However, 
since there is nearly always some transfer of 
heat from the compressor to the surrounding 
air, compression is usually polytropic rather 
than isentropic. For an air-cooled compressor, 
the transfer of heat to the air will be slight and, 
therefore, the value of the polytropic com- 
pression exponent, n, will very nearly approach 
the isentropic compression exponent, k. Hence, 
the assumption of isentropic compression for 



the ideal cycle is ordinarily not too much in 
error for an air-cooled compressor. 

Water-jacketing of the compressor cylinder 
results in lowering the temperature of the 
cylinder walls, and cooling of the vapor during 
compression will be greater for the compressor 
having a water jacket. Too, cylinder heating is 
reduced and the vapor is discharged from the 
compressor at a lower temperature. All of this 
has the effect of reducing the work of the 
compression cycle. However, the gain is 
usually not sufficient to warrant the use of a 
water jacket on most compressors, particularly 
compressors designed for R-12. For the most 
part, water-jacketing of the compressor is 
limited to compressors designed for use with 
refrigerants which have unusually high dis- 
charge temperatures, such as ammonia. Even 
then, the purpose of the jacketing is not so 
much for increasing compressor efficiency as 
it is to reduce the rate of oil carbonization and 
the formation of acids, both of which increase 
rapidly as the discharge temperature increases. 
12-26. Wet Compression. Wet compression 
occurs when small particles of unvaporized 
liquid are entrained in the suction vapor 
entering the compressor. However, theoretical 
computations indicate that wet compression 
will bring about desirable gains in compression 
efficiency and reduce the work of compression. 
This would be true if the small particles of liquid 
vaporized during the actual compression of the 
vapor. However, in actual practice, this is not 
the case. Since heat transfer is a function of 
time and since compression of the vapor in a 
modern high-speed compressor takes place very 




Volume 
Fig. 12-9. Isentropic vs. isothermal compression. 



220 



PRINCIPLES OF REFRIGERATION 



rapidly, there is not sufficient time for the liquid 
to completely vaporize during the compression 
stroke. Hence, some of the liquid particles 
remain in the vapor in the clearance volume 
and vaporize during the early part of the 
suction stroke. This action reduces the volu- 
metric efficiency of the compressor without 
benefit of the return of work to the piston by 
the expansion of the vaporizing particles. 

A result similar to this is encountered when 
excessive cooling of the cylinder reduces the 
temperature of the vapor in the clearance below 
the saturation temperature corresponding to 
the discharge. Some of the clearance vapor will 
condense and the particles of liquid formed will 
vaporize during the early part of the suction 
stroke. 

12-27. The Effect of Compressor Clearance 
on Horsepower. Theoretically, the clearance 
of the compressor has no effect on the horse- 
power, since the work done by the piston in 
compressing the clearance vapor is returned to 
the piston as the clearance vapor re-expands 
at the start of the suction stroke. However, since 
the refrigerant vapor is not an ideal gas, there 
is some loss of power in overcoming the internal 
friction of the fluid so that the power returned 
to the piston during the re-expansion of the 
clearance vapor will always be less than 
the power required to compress it. Hence, the 
clearance does have some, although probably 
slight, effect on the power requirements. 
12-28. Compressor Speed. Since the speed of 
rotation is one of the factors determining piston* 
displacement (Equation 12-1), the capacity of 
the compressor changes considerably when the 
speed of the compressor is changed. If the 
speed of the compressor is increased, the piston 
displacement is increased and the compressor 
displaces a greater volume of vapor per unit of 
time. Theoretically, based on the. assumption 
that the volumetric efficiency of the compressor 
remains constant, the capacity of the compressor 
varies in direct proportion to the speed change. 
That is, if the speed of the compressor is 
doubled, the piston displacement and capacity 
of the compressor are also doubled. Likewise, 
if the speed of the compressor is reduced, the 
piston displacement and capacity of the 
compressor are reduced in the same propor- 
tion. However, the volumetric efficiency of the 
compressor does not remain constant during 



speed changes, and therefore the change in 
compressor capacity will not be proportional 
to the speed change. 

The variation in the volumetric efficiency 
with changes in the speed of rotation is brought 
about principally by changes in the -effects of 
wiredrawing, cylinder heating, and the back 
leakage of gas through the suction and discharge 
valves. 

The amount of back leakage through the 
valves in percent per cubic foot of vapor dis- 
placed is at a maximum at low compressor 
speeds and decreases as the speed of the com- 
pressor is increased. Cylinder heating, too, is 
greatest at low compressor speeds. On the 
other hand, the effect of wiredrawing is at a 
minimum at low speeds and increases as the 
speed increases because of the increase in the 
velocity of the vapor passing through the valves. 
Hence, as the speed of rotation increases, the 
volumetric efficiency of the compressor due to 
the cylinder heating and valve leakage factors 
increases, while, at the same time, the volumetric 
efficiency due to the wiredrawing factor de- 
creases. It follows, then, that there is one 
critical speed of rotation at which the combined 
effect of these factors are at a minimum and the 
volumetric efficiency is at a maximum. From 
an efficiency standpoint, this is the speed at 
which the compressor should be operated. At 
speeds higher than this critical speed, the 
volumetric efficiency of the compressor diminish 
because the loss of efficiency due to the wire- 
drawing effect will be greater than the gain 
resulting from the decrease in the effect of 
cylinder heating and valve leakage. Likewise, at 
speeds below the critical speed, the volumetric 
efficiency will be lower because the losses 
accruing from the increase in cylinder heating 
and valve leakage will be greater than the gain 
resulting from the decrease in the wiredrawing 
losses. 

The critical speed will vary with the design of 
the compressor and with the refrigerant used, 
and can best be determined by actual test of the 
compressor. 

It is general practice in the design of modern 
high-speed compressors to use large valve ports 
in order to reduce the wiredrawing effect to a 
practical minimum. These large openings in the 
valve-plate tend to increase the clearance 
volume and decrease the volumetric efficiency 



PERFORMANCE OF RECIPROCATING COMPRESSORS 221 



due to the clearance factor, but the advantages 
accruing from the reduction in the wiredrawing 
effect more than offsets the loss of efficiency due 
to the greater clearance. This is particularly 
true where the power requirements are con- 
cerned, since the loss of power due to wire- 
drawing is much greater than the loss of power 
due to the clearance factor. 
12-29. Mechanical Efficiency. The mechanical 
friction in the compressor varies with the speed 
of rotation, but for any one speed, the mechan- 
ical friction, and therefore the friction horse- 
power, will remain practically the same at all 
operating conditions. Since the friction horse- 
power remains the same, it follows that the 
mechanical efficiency of the compressor depends 
entirely upon the loading of the compressor. 
As the total brake horsepower of the compressor 
increases due to loading of the compressor, the 
friction horsepower, being constant, will become 
a smaller and smaller percentage of the total 
horsepower .and the mechanical efficiency will 
increase. It is evident that the mechanical 
efficiency of the compressor will be greatest 
when the compressor is fully loaded. The 
mechanical efficiency of the compressor will 
vary with the design of the compressor and 
with compressor speed. An average compressor 
of good design operating fully loaded at a 
standard speed should have a mechanical 
efficiency somewhat above 90%. 
12-30. The Effect of Suction Superheat on 
Compressor Performance. It has been shown 
that superheating of the suction vapor causes 
the vapor to reach the compressor in an 
expanded condition. Therefore, when the 
vapor reaches the compressor in a superheated 
condition, the weight of refrigerant circulated 
by the compressor per minute is less than when 
the vapor reaches the compressor saturated. 
Whether or not the reduction in the weight of 
refrigerant circulated by the compressor reduces 
the refrigerating capacity of the compressor 
depends upon whether or not the superheating 
produces useful cooling. When the super- 
heating produces useful cooling, the gain in 
refrigerating capacity resulting from the increase 
in the refrigerating effect per pound is usually 
sufficient to offset the loss in refrigerating 
capacity resulting from the reduction in the 
weight of refrigerant circulated. On the other 
hand, when the superheating produces no 



useful cooling, there is no offsetting gain in 
capacity and the refrigerating capacity of the 
compressor is reduced in inverse proportion to 
the increase in the specific volume of the 
suction vapor at the compressor inlet. 

Regardless of whether or not the superheating 
produces useful cooling, the horsepower 
required to drive any one compressor is 
practically the same for a superheated cycle as 
for a saturated cycle. It was shown in Section 
8-4 that, when superheating of the vapor 
produces useful cooling, both the horsepower 
required per ton and the refrigerating capacity 
of the compressor are the same for the super- 
heated cycle as for the saturated cycle. It 
follows, then, that the horsepower required by 
any one compressor will be the same for both 
cycles. On the other hand, when superheating 
of the vapor produces no useful cooling, the 
horsepower per ton is greater than for the 
saturated cycle. However, at the same time, 
the refrigerating capacity of the compressor is 
less for the superheated cycle and the increase 
in the horsepower required per ton is more or 
less offset by the reduction in compressor 
capacity, so that the horsepower required by the 
compressor is still approximately the same as 
for the saturated cycle. Notice that, although 
the horsepower required by any one compressor 
is not appreciably changed by the superheating 
of the suction vapor, when the superheating 
does not produce useful cooling, the refriger- 
ating capacity and efficiency of the compressor 
are materially reduced. This is particularly 
true when the compressor is operating at a low 
suction temperature. It should be noted also 
that superheating of the suction vapor reduces 
the amount of cylinder heating and the efficiency 
of the compressor is increased to some extent. 
12-31. The Effect of Subcooling on Com- 
pressor Performance. When subcooling of 
the liquid refrigerant is accomplished in such a 
way that the heat given up by the liquid leaves 
the system, the specific volume of the suction 
vapor at the compressor inlet is unaffected by 
the subcooling and the weight of refrigerant 
circulated per minute by the compressor is the 
same as when no subcooling takes place. Since 
the refrigerating effect per pound is increased by 
the subcooling, the capacity of the compressor is 
increased by an amount equal to the amount of 
subcooling. Notice that the increase in the 



222 PRINCIPLES OF REFRIGERATION 



refrigerating capacity of the compressor result- 
ing from the subcooling is accomplished without 
increasing the power requirements of the com- 
pressor. Therefore, subcooling improves com- 
pressor efficiency, provided the heat given up 
during the subcooling leaves the system. 

When the heat given up during the subcooling 
does not leave the system, as when a heat 
exchanger is used, the gain in capacity due to 
the subcooling is approximately equal to the 
loss in capacity due to the superheating, and 
the refrigerating capacity of the compressor is 
very little affected. However, there is some 
small gain in compressor efficiency, since the 
superheating of the vapor in the heat exchanger 
reduces the effect of cylinder heating. 
12-32. Compressor Rating and Selection. 
As previously stated, mathematical evaluation 
of all the factors which influence compressor 
performance is not practical. Hence, com- 
pressor capacity and horsepower requirements 
are determined accurately only by actual 
testing of the compressor. Table R-10A is a 
typical compressor rating table supplied by the 
compressor manufacturer for use in compressor 
selection. The ratings have been determined by 
actual testing of the compressor under operating 
conditions set forth in the compressor testing 
and rating standards of the American Society 
of Heating, Refrigerating, and Air Conditioning 
Engineers (see Table R-10B). 

It has been shown in the foregoing sections 
that both the refrigerating capacity and the 
horsepower requirements of a compressor vary 
with the condition of the refrigerant vapor 
entering and leaving the compressor. Notice 
in Table R-10A that compressor refrigerating 
capacities (Btu/hr) and horsepower require- 
ments are listed for various saturated suction 
and discharge temperatures. The saturated 
suction temperature is the saturation tempera- 
ture corresponding to the pressure of the vapor 
at the suction inlet of the compressor, and the 
saturated discharge temperature is the saturation 
temperature corresponding to the pressure of 
the vapor at the discharge of the compressor. 

Although compressor ratings are based on 
the saturated suction and discharge tempera- 
tures, ASHRAE test standards require a certain 
amount of suction superheat and specify that 
the actual temperature of the suction vapor 
entering the compressor be those listed in 



Table R-10C. Since the compressor ratings 
given in Table R-10A are in accordance with 
ASHRAE standards, it follows that in order to 
realize the listed ratings, the suction vapor must 
enter the suction inlet of the compressor at the 
conditions shown in Table R-10C. For example, 
for a compressor operating at a saturated 
suction of —40° F, the suction vapor should 
enter the compressor at a temperature of 35° F 
(superheated 75° F from -40° F to 35° F), if 
the listed rating is to be obtained. Likewise, for 
a compressor operating at a saturated suction 
of 40° F, the actual temperature of the suction 
vapor entering the compressor should be 65° F 
(superheated 25° F from 40° F to 65" F). 
Where the actual temperature of the suction 
vapor is less than that indicated in Table 
R-10C, the tabulated rating is corrected by 
using an appropriate multiplier to obtain the 
actual compressor capacity. The multipliers 
given in Table R-10D correct the ratings to a 
basis of no superheat for the saturated condition 
listed. Where the actual suction vapor tem- 
perature is intermediate between saturation and 
the temperature shown in Table R-10C, the 
multiplier is corrected accordingly. 

The superheating is assumed to occur in the 
evaporator, in the suction line inside the re- 
frigerated space, or in a liquid-suction heat 
exchanger so that the superheat produces useful 
cooling (Section 8-4). Superheating which 
occurs outside the refrigerated space should 
be disregarded with respect to the tabulated 
ratings. 

The superheating requirement of the 
ASHRAE standards at first appears to compli- 
cate unnecessarily the compressor rating and 
selection procedure. However, this is not the 
case. The superheating requirement is very 
realistic in that the amount of superheating 
specified in the rating standards very nearly 
approaches that amount which would normally 
be expected in a well-designed application. 
Hence, the effect of the superheating require- 
ment is to cause the compressor to be rated 
under conditions similar to those under which 
the compressor will be operating in the field. 
For this reason, except in unusual cases, no 
appreciable error will occur if the compressor 
ratings given in Table R-10A are used without 
correction of any kind. Furthermore, com- 
pressor capacity requirements are not usually 



PERFORMANCE OF RECIPROCATING COMPRESSORS 223 



critical within certain limits. There are several 
reasons for this. First of all, the methods of 
determining the required compressor capacity 
(cooling load calculations) are not in themselves 
exact. Too, it is seldom possible to select a 
compressor which has exactly the required 
capacity at the design conditions. Another 
reason that compressor capacity is not critical 
within reasonable limits is that the operating 
conditions of the system do not remain constant 
at all times, but vary from time to time with the 
loading of the system, the temperature of the 
condensing medium, etc. General practice is to 
select a compressor having a capacity equal to 
or somewhat in excess of the required capacity 
at the design operating conditions. 

It was shown in Chapter 8 that subcooling of 
the liquid increases the refrigerating effect per 
pound and thereby increases compressor 
capacity. With regard to subcooling, the 
ratings given in Table R-10A are based on 
saturated liquid approaching the refrigerant 
control, that is, no subcooling. Where the 
liquid is subcooled by external means (Section 
8-7), the capacity of the compressor may be 
increased approximately 2% for each 5°F of 
subcooling. Here again, for reasons outlined in 
the preceding paragraph, the effect of subcooling 
is usually neglected in selecting the compressor. 

To select a compressor for a given application, 
the following data are needed: 

1. The required refrigerating capacity 
(Btu/hr) 

2. The design saturated suction temperature 

3. The design saturated discharge temperature 

Naturally, the required refrigerating capacity 
is the average hourly load as determined by the 
cooling load calculations. However, if an 
evaporator selection is made prior to the 
compressor selection, the compressor should be 
selected to match the evaporator capacity rather 
than the calculated load. The reasons for this 
are discussed in Chapter 13. 

The design saturated suction temperature 
depends upon the design conditions of the 
application. Specifically, it depends upon the 
evaporator temperature (the saturation tem- 
perature of the refrigerant at the evaporator 
outlet) and upon the pressure loss in the suction 
line. For instance, assume an evaporator 
temperature of 28° F and a suction line pressure 



loss of approximately 3 psi. From Table 16-3, 
the saturation pressure of Refrigerant- 12 corre- 
sponding to a temperature of 28° F is 41 .59 psia. 
Allowing for the 3 psi pressure loss in the 
suction line, the pressure of the vapor at the 
suction inlet of the compressor is 38.39 psia 
(41.59 - 3). From Table 16-3, the saturation 
temperature corresponding to a pressure of 
38.59 psia, and, therefore, the saturated suction 
temperature, is approximately 24° F. 

The design saturated discharge temperature 
depends primarily on the size of the condenser 
selected and upon the quantity and temperature 
of the available condensing medium. Methods of 
condenser selection are discussed in Chapter 14. 
12-33. Condensing Unit Rating and Selec- 
tion. Since condensing unit capacity depends 
upon the capacity of the compressor, methods 
of rating and selecting condensing units are 
practically the same as those for rating and 
selecting compressors. The only difference is 
that, whereas compressor capacities are based 
on the saturated suction and discharge tempera- 
tures, condensing unit capacities are based on 
the saturated suction temperature and on the 
quantity and temperature of the condensing 
medium. Since the size of the condenser is 
fixed at the time of manufacture for any given 
condenser loading, the only variables deter- 
mining the saturation temperature at the 
discharge of the compressor (and therefore the 
capacity of the compressor at any given suction 
temperature) is the quantity and temperature of 
the condensing medium. For air-cooled con- 
densing units, when the quantity of the air 
passing over the condenser is fixed by the fan 
selection at the time of manufacture, the only 
variable determining the capacity of the con- 
densing unit, other than the suction temperature, 
is the ambient air temperature (temperature of 
the air entering the condenser). Hence, ratings 
for air-cooled condensing units are based on 
the saturated suction temperature and the 
ambient air temperature. 

Ratings for water-cooled condensing units 
are based on the saturated suction temperature 
and on the entering and leaving water tempera- 
tures.* Typical capacity ratings for air-cooled 

* For any given condenser loading and entering 
water temperature, the leaving water temperature 
depends only on the quantity of water (gallons per 
minute) flowing through the condenser. 



224 PRINCIPLES OF REFRIGERATION 



and water-cooled condensing units are given in 
Tables R-ll and R-12, respectively. 

Example 12-12. A certain refrigeration 
application has a calculated cooling load of 
33,000 Btu/hr. If the design saturated suction 
and discharge temperatures are 20° F and 
100° F, respectively, select a compressor from 
Table R-10A which will meet the requirements 
of the application. 

Solution. Locate the desired saturated dis- 
charge temperature in the first column of the 
table (100° F). Next, in the second column, 
locate the desired saturated suction tempera- 
ture and read to the right until a compressor 
having a capacity equal to or somewhat in 
excess of the desired capacity is found. Select 
compressor, Model #5F20, which has a capacity 
of 34,000 Btu/hr at 1450 rpm. 

Example 12-13. Assume that the design 
saturated suction in Example 12-12 is 0° F and 
make a new compressor selection. 

Solution. Using the procedure outlined in the 
solution to Example 12-12, select compressor 
Model #5F30 which has a capacity of 36,000 
Btu/hr at 1750 rpm. 

Example 12-14. Determine the capacity of 
compressor Model #5F20 when operating at a 
23° F saturated suction temperature, if the 
saturated discharge temperature is 100° F. 



Approximate capacity 
of compressor at a 23° F 
saturated suction 



= 34,000 + 2780 
= 37,380 Btu/hr 



Solution. From Table R-10A, 


Compressor capacity 
at 30° F suction 


= 44,200 Btu/hr 


Compressor capacity 
at 20° F suction 


= 34,600 Btu/hr 


Capacity change per 
10° F change in saturated 
suction temperature 

Average capacity change 
per ° F change in suction 
temperature 


= 44,200 - 34,600 
= 9600 Btu/hr 

9600 
10 


Total capacity change 
for 3° F change in suction 
temperature 


= 960 Btu/hr 

= 960 x 3 
= 2780 Btu/hr 



Example 12-15. A certain refrigeration 
application has a calculated cooling load of 
8750 Btu/hr and the ambient temperature is 
90° F. If the design saturated suction tempera- 
ture is 20° F, select an air-cooled condensing 
unit which will satisfy the requirements of the 
application. 

Solution. From Table R-ll, select a 1 hp 
condensing unit having a capacity of 9340 
Btu/hr at the prescribed conditions. 

PROBLEMS 

1. A four-cylinder reciprocating compressor 
having a 2 in. bore and a 2.5 in. stroke is 
rotating at 100 rpm. Compute the piston 
displacement in cubic feet per minute. 

Arts. 18.18 cfm 

2. Assume the compressor in Problem 12-1 is 
operating on the cycle described in Problem 7-1 
and compute the theoretical refrigerating 
capacity of the compressor in Btu/hr. 

Ans. 2.47 tons 

3. Using the conditions of Problem 12-2, 
determine: 

(a) The volumetric efficiency of the com- 
pressor. 

(b) The actual refrigerating capacity of the 
compressor in tons. Ans. 1.52 tons 

(c) The brake horsepower required per ton 
(allow 10% for mechanical friction and 
assume the compression efficiency is the 
same as the volumetric efficiency). 

Ans. 2.7 hp/ton 

(d) The total brake horsepower required to 
drive the compressor. Ans. 4.1 hp 

4. From the compressor rating tables, select a 
compressor which will satisfy the following 
conditions: 

(a) Required capacity 24,900 Btu/hr 

(b) Design saturated suction temperature 
10° F 

(c) Design saturated discharge temperature 
105° F 



13 

System Equilibrium 
and Cycling 
Controls 



13-1. System Balance. In the designing of a 
refrigerating system, one of the most important 
considerations is that of establishing the proper 
relationship or "balance" between the vapor- 
izing and condensing sections of the system. 
It is important to recognize that whenever an 
evaporator and a condensing unit are connected 
together in a common system, a condition 
of equilibrium or "balance" is automatically 
established between the two such that the rate 
of vaporization is always equal to the rate of 
condensation. That is, the rate at which the 
vapor is removed from the evaporator and 
condensed by the condensing unit is always 
equal to the rate at which the vapor is produced 
in the evaporator by the boiling action of the 
liquid refrigerant. Since all the components in 
a refrigerating system are connected together in 
series, the refrigerant flow rate through all 
components is the same. It follows, therefore, 
that the capacity of all the components must be 
of necessity the same. 

Obviously, then, where the system com- 
ponents are selected to have equal capacities at 
the system design conditions, the point of 
system equilibrium or balance will occur at the 
system design conditions. On the other hand, 
when the components selected do not have 
equal capacities at the system design conditions, 
system equilibrium will be established at 
operating conditions other than the system 



design conditions and the system will not 
perform satisfactorily. 

In any event, it is important to understand 
that, regardless of the equipment selected, the 
system will always establish equilibrium at some 
conditions such that all the system components 
will have equal capacity. Hence, whether or not 
system equilibrium is established at the sys- 
tem design conditions depends entirely upon 
whether or not the equipment is selected to have 
approximately equal capacities at the system 
design conditions. This concept is best illu- 
strated through the use of a series of examples. 

Example 13-1. A walk-in cooler, having a 
calculated cooling load of 11,000 Btu/hr, is to 
be maintained at 35° F. The desired evaporator 
TD is 12° F and the ambient temperature is 
90° F. Allowing 3° F (equivalent to approxi- 
mately 2 lb) for the pressure drop in the suction 
line (see Section 12-32, select an air-cooled 
condensing unit and a unit cooler from manu- 
facturer's catalog data. 

Solution. Since the design space temperature 
is 35° F and the design evaporator TD is 12° F, 
the design evaporator temperature is 23° F 
(35° F - 12° F). When we allow for a 3° F 
loss in the suction line resulting from pressure 
drop, the saturation temperature at the com- 
pressor suction is 20° F (23° F - 3° F). 

From Table R-ll, select lj hp condensing 
unit, which has a capacity of 12,630 Btu/hr at a 
20° F saturated suction and 90° F ambient air 
temperature. Although the condensing unit 
capacity is somewhat in excess of the calculated 
load of 11,000 Btu/hr, it is sufficiently close to 
make the condensing unit acceptable. How- 
ever, to assure proper system balance, the unit 
cooler selection must now be based on the 
condensing unit capacity of 12,630 Btu/hr 
rather than on the calculated load of 11,000 
Btu/hr. * Hence, a unit cooler having a capacity 
of approximately 12,630 Btu/hr at a 12° F TD 
is required. 

From Table R-8, unit cooler Model #105 has 
a capacity of 10,500 Btu/hr at a 10° F TD. 
Using the procedure outlined in Section 11-23, 
it can be determined that this unit cooler will 
have a capacity of 1 2,600 Btu/hr when operating 

* Either the evaporator or the condensing unit 
may be selected first. However, once either one has 
been selected, the other must be selected for approxi- 
mately the same capacity. 



225 



226 PRINCIPLES OF REFRIGERATION 































































Nf 


























^& 


























(Compressor; - 




















\r 


fc 










\*w 










\\ 






















V 


\ 
























1 > 



0' 



5° 10° 15° 20° 
Suction temperature 

J I I I I L 



25° 



32° 



V 



27° 22° XT 12° 
Evaporator TD 
Fig. 13-1. Graphic analysis of system balance. 



at a 12° F TD. Since this is very close to the 
condensing unit capacity, the unit cooler is 
ideally suited for the application. 
13-2. Graphical Analysis of System Equi- 
librium. For any particular evaporator and 
condensing unit connected together in a 
common system, the relationship established 
between the two, that is, the point of system 
balance, can be evaluated graphically by 
plotting evaporator capacity against condensing 
unit capacity on a common graph. Using data 
taken from the manufacturers' rating tables, 
condensing unit capacity is plotted against 
suction temperature, whereas evaporator capa- 
city is plotted against evaporator TD. A 
graphical analysis of the system described in 
Example 13-1 is shown in Fig. 13-1. 

In order to understand the graphical analysis 
of system equilibrium in Fig. 1 3-1 , it is important 
to recognize that, for any given space tempera- 
ture, there is a fixed relationship between the 
evaporator TD and the compressor suction 
temperature. That is, for any given space 
temperature, once the evaporator TD is 
selected, there is only one possible suction 
temperature which will satisfy the design 
conditions of the system. 

Notice that, in Example 13-1, for the design 
space temperature of 35° F and assuming a 
3° F loss in the suction line, the only possible 
suction temperature that can coexist with the 
design evaporator TD of 12° F is 20° F. In this 



instance, the evaporator TD will be 12° F when, 
and only when, the suction temperature is 20° F. 
Any suction temperature other than 20° F will 
result in an evaporator TD either greater or 
smaller than 12° F. For example, assume a 
suction temperature of 25° F. Adding 3° F to 
allow for the suction line loss, the evaporator 
temperature is found to be 28° F (25° F + 3° F). 
Then subtracting the evaporator temperature 
from the space temperature, it is determined 
that the evaporator TD will be 7° F (35° F - 
28° F) when the suction temperature is 25° F. 
Using the same procedure, it can be shown that 
if the suction temperature is reduced to 15° F, 
the evaporator TD will increase to 17° F, and 
when the suction temperature is 10° F, the 
evaporator TD will be 22° F, and so on. 

Apparently, then, raising or lowering the 
suction temperature always brings about a 
corresponding adjustment in the evaporator TD. 
Provided that the space temperature is kept 
constant, raising the suction temperature 
reduces the evaporator TD, whereas lowering 
the suction temperature increases the evaporator 
TD. 

With regard to Fig. 13-1, the following 
procedure is used in making a graphical analysis 
of the system equilibrium conditions: 

1. On graph paper, lay out suitable scales 
for capacity (Btu/hr), suction temperature (° F), 
and evaporator TD (° F). The horizontal lines 
are used to represent capacity, whereas the 
vertical lines are given dual values, representing 
both suction temperature and evaporator TD. 
The latter is meaningful, however, only when 
the suction temperature and evaporator TD 
scales are so correlated that the two conditions 
which identify any one vertical line are condi- 
tions which, at the design space temperature, 
actually represent conditions that will occur 
simultaneously in the system. The procedure 
for correlating the suction temperature and 
evaporator TD scales was discussed in the 
preceding paragraphs. 

2. Using manufacturer's catalog data, plot 
the capacity curve for the condensing unit. 
Since condensing unit capacity is not exactly 
proportional to suction temperature, the con- 
densing unit capacity curve will ordinarily have 
a slight curvature. Hence, for accuracy, a 
capacity point is plotted for each of the suction 



SYSTEM EQUILIBRIUM AND CYCLING CONTROLS 227 



temperatures listed in the table and these points 
are connected with the "best-fitting" curve. 

3. From the evaporator manufacturer's cata- 
log data, plot the evaporator capacity curve. 
Since evaporator capacity is assumed to be 
proportional to the evaporator TD, the evapo- 
rator capacity curve isa straight line, the position 
and direction of which is adequately established 
by plotting the evaporator capacity at any two 
selected TDs. The evaporator capacity at any 
other TD will fall somewhere along a straight 
line drawn through these two points. In Fig. 
13-1, evaporator TDs of 7°F and 12° F are 
used in plotting the two capacity points required 
to establish the evaporator capacity curve. 

Notice in Fig. 13-1, that as the suction 
temperature increases, the evaporator TD 
decreases. This means, in effect, that as the 
suction temperature increases the capacity of 
the evaporator decreases while the capacity of 
the condensing unit (compressor) increases. 
Likewise, as the suction temperature decreases, 
the capacity of the evaporator increases while 
the capacity of the condensing unit decreases. 
The intersection of the two capacity curves 
indicates the point of system equilibrium. In 
this instance, because the evaporator and 
condensing unit have been selected to have 
equal capacities at the system design conditions, 
the point of system equilibrium occurs at the 
system design conditions (12° F TD and 20° F 
suction temperature). Although the total 
system capacity is somewhat greater than the 
calculated load, the difference is not sufficiently 
great to be of any particular consequence, and 
means only that the system will operate fewer 
hours out of each 24 than was originally 
anticipated.* The relationship between system 
capacity and the calculated load is discussed 
more fully later in the chapter. 

It has already been pointed out that where the 
evaporator and condensing unit selected do not 
have equal capacities at the system design 
conditions, the point of system equilibrium will 

* For simplicity, the heat given off by the evapo- 
rator fan motor has been neglected. If this heat is 
added to the cooling load, the total system capacity 
would be almost exactly equal to the calculated load. 
It is not often that equipment can be found which so 
nearly meets the requirements of an application as in 
this instance. 



occur at conditions other than the design 
conditions. For instance, assume that unit 
cooler Model #UC-120, rather than Model 
#UC-105, is selected in the foregoing example. 
At the design evaporator TD of 12° F, this unit 
cooler has capacity of 14,000 Btu/hr, whereas 
the condensing unit selected has a capacity of 
only 12,630 Btu/hr at the design suction 
temperature of 20° F. Consequently, at the 
design conditions, evaporator capacity will be 
greater than condensing unit capacity, that is, 
vapor will be produced in the evaporator at a 
greater rate than it is removed from the evapo- 
rator and condensed by the condensing unit. 
Therefore, the system will not be in equilibrium 
at these conditions. Rather, the excess vapor 
will accumulate in the evaporator and cause an 
increase in the evaporator temperature and 
pressure. Since raising the evaporator tempera- 
ture increases the suction temperature and, at 
the same time, reduces the evaporator TD, the 
condensing unit capacity will increase and the 
evaporator capacity will decrease. System 
equilibrium will be established when the 
evaporator temperature rises to some point 
where the suction temperature and evaporator 
TD are such that the condensing unit capacity 
and the evaporator are equal. In this instance, 
system equilibrium, as determined graphically 
(point A in Fig. 13-2), is established at a 
suction temperature of approximately 23° F. 
The evaporator TD is approximately 9° F, 
which is 3° F less than the design evaporator 
TD of 12° F and which will result in a space 
humidity somewhat higher than the design 
condition. The total system capacity is approxi- 
mately 13,500 Btu/hr, which is about 23% 
greater than the calculated hourly load of 
11,000 Btu/hr. This means that the system 
running time will be considerably shorter than 
originally calculated. For instance, if the 
original load calculation is based on a 16-hr 
running time, the system will operate only 
about 13 hr out of each 24. 

The question immediately arises as to whether 
or not this system will perform satisfactorily. 
Although this would depend somewhat on the 
particular application, the answer is that it 
probably would not in the majority of cases. 
There are several reasons for this. First, the 
evaporator TD of 9° F is considerably less than 
the design TD of 12° F and would probably 



228 PRINCIPLES OF REFRIGERATION 



28 

26 

24 

| 22 

§20 

| 18 

8 16 

I"l4 

&12 

10 

8 













































\ 
























X 


<"' 




















^\B 






















' 


1 1 \ 
























\ 


























\ 






^p 


II 


z 




— 


™ 






— 






I — 

\ 
























^r 










l\ 


\\ 






\ 










1 \ 

i 


1 

1 


0* 5* 10* 15* 20* 25' 
Suction temperature 

1 1 1 1 1 1 



32* 27* 



7* 



22* 17* 12* 
Evaporator TD 

Fig. 13-Z 



result in a space humidity too high for the appli- 
cation. Ordinarily, the humidity in the refrig- 
erated space must be maintained within certain 
fixed limits. Assuming that the design TD is 
selected to produce the median condition within 
these limits, a one degree deviation from the 
design TD in either direction is usually the 
maximum which can be allowed if the space 
humidity is to be maintained within the limits 
specified for the application. 

Another consideration is the fact that the 
system capacity is some 23% greater than the 
calculated load so that the system operating 
time will be relatively short. Although the sys- 
tem capacity exceeds the calculated load by a 
larger margin than good practice prescribes, this 
in itself would not ordinarily cause any serious 
problem in a majority of applications. How- 
ever, since the shorter running time will also 
tend to aggravate the already existing problem 
of high humidity, especially in the wintertime, 
when the two conditions are taken together, it 
seems unlikely that the system would produce 
satisfactory results in any application where the 
space humidity is an important factor. 

In the event that the equipment in question 
represents the best available selection, the ques- 
tion arises as to what can be done to bring the 
system into balance at conditions more in 
keeping with the design conditions. In this 
instance, since the problem is one of excessive 
evaporator capacity with relation to the con- 
densing unit capacity at the design conditions, 



logical corrective measures prescribe either an 
increase in the condensing unit capacity or a 
reduction in the evaporator capacity in order to 
re-establish the point of system equilibrium at 
conditions nearer to the design conditions. 

Which of these two measures will produce the 
most satisfactory results depends upon the rela- 
tionship between the over-all system capacity 
and the calculated load. Whereas increasing 
either the condensing unit capacity or the evapo- 
rator capacity will always bring about an in- 
crease in the over-all system capacity, reducing 
either the condensing unit capacity or the 
evaporator capacity will always bring about a 
reduction in the over-all system capacity. 
Referring to Fig. 13-2 for the system under con- 
sideration, if the condensing unit capacity is 
increased to the evaporator capacity at the 
design conditions, system equilibrium will shift 
from point A to point B. On the other hand, if 
the evaporator capacity is reduced to the con- 
densing unit capacity at the design conditions, 
the point of system equilibrium will shift from 
A to C. Notice that, although the system is 
balanced at the design conditions at either 
points B or C, the over-all system capacity at 
point B is considerably above the calculated 
load, whereas at point C the over-all system 
capacity very nearly approaches the calculated 
load. Hence, in this instance, it is evident that 
increasing the condensing unit capacity as a 
means of bringing the system into balance at the 
design conditions cannot be recommended, 
since it would also increase the over-all system 
capacity and therefore tend to aggravate the 
already existing problem of excessive system 
capacity with relation to the calculated load. 
On the other hand, in addition to bringing the 
system into balance at the design conditions, 
reducing the evaporator capacity will also have 
the beneficial effect of reducing the over-all 
system capacity and thereby bringing it more 
into line with the calculated load. 
13-3. Decreasing or Increasing Evaporator 
Capacity. Reducing the evaporator capacity 
can be accomplished in several ways. One is to 
"starve" the evaporator, that is, to reduce the 
amount of liquid refrigerant in the evaporator 
by adjusting the refrigerant flow control so that 
the evaporator is only partially flooded with 
liquid. This effectively reduces the size of the 
evaporator, since that part of the evaporator 



SYSTEM EQUILIBRIUM AND CYCLING CONTROLS 229 



which is not filled with liquid becomes, in effect, 
a part of the suction line. 

Another method of reducing the capacity of 
the evaporator is to reduce the air velocity over 
the evaporator by slowing the evaporator fan or 
blower. However, this method has its limita- 
tions in that the air velocity must be maintained 
at a level sufficient to assure adequate air 
circulation in the refrigerated space. Too, re- 
ducing the air quantity causes a change in the 
sensible heat ratio of the evaporator. Depending 
upon the particular application, this may or may 
not be desirable. 

As a general rule, there is little, if anything, 
that can be done to increase the capacity of an 
undersized evaporator. Occasionally, the evapo- 
rator capacity can be increased by increasing 
the air quantity. However, since increasing the 
air quantity also increases air velocity and fan 
horsepower requirements, this method has its 
limitations for reasons already discussed in 
Section 11-22. 

In some cases, the evaporator surface area 
can be increased somewhat by using a length of 
eitherjjare tubing or finned tubing as a "drier 
loop" or as additional evaporator surface. How- 
ever, this too has its limitations because of the 
pressure drop accruing in the tubing. 
13-4. Decreasing or Increasing Condensing 
Unit Capacity. Decreasing the condensing 
unit capacity can be accomplished in several 
ways, all of which involve decreasing the com- 
pressor displacement. Probably the simplest 
and most common method of reducing the con- 
densing unit capacity is to reduce the speed of 
the compressor by reducing the size of the pulley 
on the compressor driver. The speed reduction 
required is approximately proportional to the 
desired capacity reduction. 

The relationship between the speed of the 
compressor and the speed of the compressor 
driver is expressed in the following equation: 

Rpm 1 x D x = Rpm t x D 2 (13-1) 

where Rpm x = the speed of the compressor 
(rpm) 
D 1 = the diameter of the compressor 
flywheel (inches) 
Rpm t = the speed of the compressor 
driver (rpm) 
Z) 2 = the diameter of the driver pulley 
(inches) 



Note. Where the compressor driver is a 
four-pole, alternating-current motor operating 
on 60 cycle power, the approximate driver speed 
is 1750 rpm. For a two-pole, alternating-current 
motor, the approximate speed is 3500 rpm. 

Example 13-2. A refrigeration compressor 
having a 10 in. flywheel is driven by a four-pole, 
alternating-current motor. If the diameter of 
the motor pulley is 4 in., determine the speed 
of the compressor. 

Solution. Rearranging and 
applying Equation 13-1, R x D 

Rp mi = ^D, 

_ 1750 x 4 

10 
= 700 

Example 13-3. Determine the diameter of 
the motor pulley required to reduce the speed 
of the compressor in Example 13-2 from 700 to 
600 rpm. 



Solution. Rearranging and 
applying Equation 13-1, D s 



_ Rpm J x D t 
Rpm^ 
600 x 10 
1750 
= 3.5 in. 

Another method of reducing condensing unit 
capacity is to reduce the volumetric efficiency of 
the compressor by increasing the clearance 
volume. This increase is accomplished by in- 
stalling a thicker gasket between the cylinder 
housing and the valve-plate. 

In some cases, small increases in condensing 
unit capacity can be obtained by merely in- 
creasing the speed of the compressor. However, 
when the capacity increase needed is substan- 
tial, it is usually more practical and more 
economical to use a larger size condensing unit 
and reduce the capacity as necessary. The 
reasons for this are several. 

First, since the increase in compressor capa- 
city will be accompanied by an increase in the 
horsepower requirements, any substantial in- 
crease in the compressor capacity will tend to 
overload the compressor driver and necessitate 
the use of a larger size. Too, some thought must 
be given to the condenser capacity. Here again, 
any increase in compressor capacity will tend to 
place a heavier load on the condenser. If the 
size of the condenser is not increased in propor- 
tion to the increase in the condenser load, 



230 PRINCIPLES OF REFRIGERATION 



1 



Fig. 13-3 



excessive compressor discharge temperature and 
pressure will result. Not only will this materially 
reduce the life of the equipment and increase 
maintenance and operating costs, but it will also 
tend to nullify to some extent the gain in capa- 
city originally accruing from the increase in 
compressor speed. 

It is apparent from the foregoing that, in most 
cases, increasing the capacity of either the eva- 
porator or the condensing unit is something 
which is not easily accomplished. Therefore, it 
is usually more practical and more economical 
to select oversized equipment rather than under- 
sized equipment. When the evaporator or the 
condensing unit is oversized and capacity reduc- 
tion is required to bring the system components 
into balance at the desired conditions, the capa- 
city reduction can readily be made with little, if 
any, loss in system efficiency. 
13-5. System Capacity vs. Calculated Load. 
The relationship between system capacity and 
system load is one which warrants careful con- 
sideration and which can be best explained by 
comparing the refrigerating system to a water 
pumping system. For example, assume that it 
is desired to maintain a constant water level in 
the tank shown in Fig. 13-3. If the water flows 
into the tank at a fixed and constant rate which 
is readily computable, the water in the tank can 
be maintained at a fixed level simply by installing 
a pumping system which has a capacity exactly 
equal to the flow rate of the water into the tank. 
Since the flow rate of the water entering the tank 
is constant and since the pumping rate is equal 
to the water flow rate, the pump will operate 
continuously and no other water level control of 
any kind will be needed. 



On the other hand, if the flow rate of the water 
entering the tank varies from time to time, it is 
evident that if the level of the water is to be 
maintained within fixed limits, the pumping 
system must be selected to have a capacity equal 
to or somewhat in excess of the highest sustained 
flow rate of the water entering the tank. It is 
evident also that some means of cycling the 
pump "off" and "on" must be provided. Other- 
wise, during periods when the flow rate of the 
water entering the tank is less than maximum, 
the pumping rate will be excessive and the level 
of the water in the tank will be reduced below 
the desired levei. One convenient and practical 
means of cycling the pump is to install a float 
control in the tank (Fig. 1 3-4). The float control 
is arranged to close the electrical contacts and 
start the pump when the water in the tank rises 
to a predetermined maximum level. When the 
water level in the tank falls to a predetermined 
lower limit, the float control acts to close the 
electrical contacts and stop the pump. In this 
way, intermittent operation of the pump will 
maintain a relatively constant water level in 
the tank. 

The latter principle is readily applied to the 
refrigerating system. Since the cooling load on a 
refrigerating system varies from time to time, 
the system is usually designed to have a capacity 
equal to or somewhat in excess of the average 
maximum cooling load. This is done so that the 
temperature of the space or product can be 
maintained at the desired low level even under 
peak load conditions. As in the case of the water 




To pump motor 



Fig. 13-4 



SYSTEM EQUILIBRIUM AND CYCLING CONTROLS 231 



pumping system, since the system refrigerating 
capacity will always exceed the actual cooling 
load, some means of cycling the system "off" 
and "on" is needed in order to maintain the 
temperature of the space or product at a con- 
stant level within reasonable limits and to pre- 
vent the temperature of the space or product 
from being reduced below the desirable 
minimum. 

For any refrigerating system, the relative 
length of the "off" and "on" cycles will vary 
with the loading of the system. During periods 
of peak loading, the "running" or "on" cycles 
will be long and the "off" cycles will be short, 
whereas during periods of minimum loading the 
"on" cycles will be short and the "off" cycles 
will be long.* 

13-6. Cycling Controls. The controls used to 
cycle a refrigerating system "on" and "off" are 
of two principal types: (1) temperature actuated 
(thermostatic) and (2) pressure actuated. Each 
of these types is discussed in the following 
sections. 

13-7. Temperature Actuated Controls. 
Temperature actuated controls are called ther- 



Bellows 



^m 




<a) 






Diaphragm ■~ >k . 



(b) 



Fig. 13-5. Bulb-type temperature sensing element. 

* Unlike the water pumping system, refrigerating 
systems are designed to have sufficient capacity to 
permit "off" cycles even during periods of peak 
loading. This is necessary in order to allow time for 
defrosting of the evaporator. However, allowances 
are made for defrosting time in the load calculations 
(the 24-hr load, is divided by the desired running 
time to obtain the average hourly load) and need 
not be further considered when selecting the equip- 
ment. 



Stationary 
contact 




Bellows 




Fig. 13-4. Schematic diagram of simplified pressure 
control. 



mostats. Whereas float controls are sensitive to 
and are actuated by changes in liquid level, 
thermostats are sensitive to and are actuated by 
changes in temperature. Thermostats are used 
to control the temperature level of a refrigerated 
space or product by cycling the compressor 
(starting and stopping the compressor driving 
motor) in the same way that float controls are 
used to control liquid level by cycling the pump 
(starting and stopping the pump motor). 
13-8. Temperature Sensing Elements. Two 
types of elements are commonly used in thermo- 
stats to sense and relay temperature changes to 
the electrical contacts or other actuating mecha- 
nisms. One is a fluid-filled tube or bulb which is 
connected to a bellows or diaphragm and filled 
with a gas, a liquid, or a saturated mixture of 
the two (Figs. 13-5aand 13-56).t Increasing the 
temperature of the bulb or tube increases the 
pressure of the confined fluid which acts through 
the bellows or diaphragm and a system of levers 
to close electrical contacts or to actuate other 
compensating mechanisms (Fig. 13-6). Decreas- 
ing the temperature of the tube or bulb will 
have the opposite effect. 

f The thermostat described here is called a 
remote-bulb thermostat. Although there are a 
number of different types of thermostats, this is the 
type most frequently used in commercial refriger- 
ation applications. Thermostats are used for many 
purposes other than controlling a compressor 
driving motor, as, for example, opening and closing 
valves, starting and stopping damper motors, etc. 



232 PRINCIPLES OF REFRIGERATION 



/ 



Dissimilar 
metals 



Invar- 



___ Brass 
"or steel 

_Bimetal 
"element 



Normal 
M 





Fig. 13-7. Bimetal-t/pe tem- 
perature sensing element. 



Another and entirely different temperature 
sensing element is the compound bar, commonly 
called a bimetal element. The compound bar is 
made up of two dissimilar metals (usually Invar 
and brass or Invar and steel) bonded into a flat 
strip (Fig. 13-7a). Invar is an alloy which has a 
very low coefficient of expansion, whereas brass 
and steel have relatively high coefficients of 
expansions. Since the change in the length of 
the Invar per degree of temperature change will 
always be less than that of the brass or steel, 
increasing the temperature of the bimetal ele- 
ment causes the bimetal to warp in the direction 
of the Invar (the inactive metal) as shown in 
Fig. 13-76, whereas decreasing the temperature 
of the bimetal element causes the bimetal to 
warp in the direction of the brass or steel (the 
active metal) as shown in Fig. 13-7c. This 
change in the configuration of the bimetal ele- 
ment with changes in temperature can be utilized 
directly or indirectly to open and close electrical 
contacts or to actuate other compensating 
mechanisms. 

13-9. Differential Adjustment. Like float 
controls, thermostats have definite "cut-in" and 
"cut-out" points. That is, the thermostat is 
adjusted to start the compressor when the tem- 
perature of the space or product rises to some 
predetermined maximum (the cut-in tempera- 
ture) and to stop the compressor when the 
temperature of the space or product is reduced 
to some predetermined minimum (the cut-out 
temperature). 

The difference between the cut-in and cut- 
out temperatures is called the differential. In 
general, the size of the differential depends upon 
the particular application and upon the location 



of the temperature sensing element. Where the 
temperature sensing element of the thermostat is 
located in or on the product and controls the 
product temperature directly, the differential is 
usually small (2° F or 3° F). On the other hand, 
where the sensing element is located in the space 
and controls the space temperature, the differen- 
tial is ordinarily about 6° F or 7° F. In many 
instances, the sensing element of the thermostat 
is clamped to the evaporator so that the space 
or product temperature is controlled indirectly 
by controlling the evaporator temperature, in 
which case the differential used must be larger 
(15° F to 20° F or more) in order to avoid short- 
cycling of the equipment. 

When the thermostat controls the space or 
product temperature directly, the average space 
or product temperature is approximately the 
median of the cut-in and cut-out temperatures. 
Therefore, to maintain an average space tem- 
perature of 35° F, the thermostat can be adjusted 
for a cut-in temperature of approximately 38° F 
and a cut-out temperature of approximately 
32° F. 

On the other hand, when the space tempera- 
ture is controlled indirectly by controlling the 
evaporator temperature, an allowance must be 
made in the cut-out setting to compensate for 
the evaporator TD. For example, for an 
average space temperature of 35° F and assum- 
ing an evaporator TD of 12° F, to compensate 
for the evaporator TD, the cut-out temperature 
would be set at 20° F (32° F - 12°) rather than 
at 32° F. Notice that the cut-in temperature is 
set at 38° F in either case. This is because the 
space temperature and the evaporator tempera- 
ture are the same at the time that the system 



SYSTEM EQUILIBRIUM AND CYCLING CONTROLS 233 



cycles on. After the compressor cycles off the 
evaporator continues to absorb heat from the 
space and warms up to the space temperature 
during the ofT cycle (Fig. 13-8). Therefore, when 
the space temperature rises to the cut-in tem- 
perature of 38° F, the evaporator will also be at 
the cut-in temperature of 38° F. As soon as the 
compressor is started, the evaporator tempera- 
ture is quickly reduced below the space tempera- 
ture by an amount approximately equal to the 
design evaporator TD. Therefore, in this in- 
stance, when the space temperature is reduced 
to 32° F (the desired minimum), the evaporator 
temperature (which the thermostat is controlling) 
will be approximately 20° F (12° F less than the 
space temperature). 

Regardless of whether the thermostat controls 
the space temperature directly or indirectly, 
proper adjustment of the cut-in and cut-out 
temperatures is essential to good operation. If 
the cut-in and cut-out temperatures are set too 
close together (differential too small) the system 
will have a tendency to short-cycle (start and 
stop too frequently). This will materially reduce 
the life of the equipment and may result in other 
unsatisfactory conditions. On the other hand, 
if the cut-in and cut-out temperatures are set too 
far apart (differential too large), the on and off 
cycles will be too long and unnecessarily large 
fluctuations in the average space temperature 
will result. Naturally, this too is undesirable. 

Although approximate cut-in and cut-out 
temperature settings for various types of appli- 
cations have been determined by field experience, 



in many cases it is necessary to use trial-and- 
error methods to determine the optimum settings 
for a specific installation. 
13-10. Range Adjustment. In addition to the 
differential, cycling controls have another adjust- 
ment, called the "range," which is also asso- 
ciated with the cut-in and cut-out temperatures. 
Although, like the differential, the range can be 
defined as the difference between the cut-in and 
cut-out temperatures, the two are not the same. 
For example, assume that a thermostat is 
adjusted for a cut-in temperature of 30° F and a 
cut-out temperature of 20° F. Whereas the 
differential is said to be 10° F (30° - 20°), the 
range is said to be between 30° F and 20° F. 

Although it is possible to change the range 
without changing the differential, it is not 
possible to change the differential without 
changing the range. For instance, suppose that 
the thermostat previously mentioned is re- 
adjusted so that the cut-in temperature is raised 
to 35° F and the cut-out temperature is raised to 
25° F. Although the differential is still 10° F 
(35° — 25°), the operating range of the control 
is 5° F higher than it was originally, that is, the 
operating range is now between 35° F and 25° F, 
whereas previously it was between 30° F and 
20° F. In this instance, the range of the control 
is changed, but the differential remains the same. 
Under the new control setting the average space 
temperature will be maintained approximately 
5° F higher than under the old setting. 

Suppose now that the differential is increased 
5° F by raising the cut-in temperature from 30° F 




-Average evaporator temperature 

-Minimum evaporator temperature (cut-out temperature) 



Fig. 13-8. Notice that when the unit cycles "on," the evaporator temperature is the same as the space 
temperature, whereas when the unit cycles "off," the evaporator temperature is lower than the space 
temperature by an amount equal to the design evaporator TD. 



234 PRINCIPLES OF REFRIGERATION 

, Differential adjustment 




Fig. 13-9. Schematic diagram 
of thermostatic motor control 
illustrating range and differen- 
tial adjustments. 



Range 
adjustment 



to 35° F while the cut-out temperature is left at 
the original setting of 20° F. Notice that both 
the differential and the range are changed. The 
differential, originally 10° F, is now 15° F and 
the range, originally between 30° F and 20° F, is 
now between 35° F and 20° F. With this control 
setting, the running cycle will be somewhat 
longer because the differential is larger. Too, 
the average space temperature will be 2 or 3 
degrees higher because the cut-in temperature is 
higher. If the differential had been increased by 
lowering the cut-out temperature 5° F rather 
than by raising the cut-in temperature 5° F, the 
operating range of the control would have 
shifted to the opposite direction and the average 
space temperature would have been 2 or 3 
degrees lower than the original space tempera- 
ture. 

Typical range and differential adjustments 
are shown in Fig. 13-9. Turning the range- 
adjusting screw clockwise increases the spring 
tension which the bellows pressure must over- 
come in order to close the contacts and, 
therefore, raises both the cut-in and cut-out 
temperatures. Turning the range-adjusting 
screw counterclockwise decreases the spring 
tension and lowers both the cut-in and cut- 
out temperatures. 



Turning the differential-adjusting screw clock- 
wise causes the limit bar A to move toward the 
screw head, thereby increasing the travel of the 
pin B in the slot. This has the effect of in- 
creasing the differential by lowering the cut-out 
temperature. Turning the differential adjusting 
screw counterclockwise raises the cut-out tem- 
perature and reduces the differential. By manipu- 
lating both range and differential adjustments, 
the thermostat can be adjusted for any desired 
cut-in and cut-out temperatures. 

The arrangement shown in Fig. 13-9 repre- 
sents only one of a number of methods which 
can be employed to adjust the cut-in and cut-out 
temperatures. The particular method used in 
any one control depends on the type of control 
and on the manufacturer. For example, for the 
control shown in Fig. 13-9, changing the range 
adjustment changes both the cut-in and cut-out 
temperatures simultaneously, whereas for an- 
other type of control changing the range adjust- 
ment changes only the cut-in temperature. For 
still another type of control, changing the range 
adjustment changes only the cut-out tempera- 
ture. However, whatever the method of adjust- 
ment, the principles involved are similar and the 
exact method of adjustment is readily deter- 
mined by examining the control . In many cases, 



SYSTEM EQUILIBRIUM AND CYCLING CONTROLS 235 



instructions for adjusting the control are given 
on the control itself. 

If electrical contacts are permitted to open or 
close slowly, arcing will occur between the con- 
tacts, and burning or .welding together of the 
contacts will result. Therefore, cycling controls 
which employ electrical contacts must all be 
equipped with some means of causing the con- 
tacts to open and close rapidly in order to avoid 
arcing. In Fig. 13-9, the armature and per- 
manent magnet serve this purpose. As the 
pressure in the bellows increases and the 
movable contact moves toward the stationary 
contact, the strength of the magnetic field 
between the armature and the horseshoe magnet 
increases rapidly. When the movable contact 
approaches to within a certain, predetermined, 
minimum distance of the stationary contact, the 
strength of the magnetic field becomes great 
enough to overcome the opposing spring tension 
so that the armature is pulled into the magnet 
and the contacts are closed rapidly with a snap 
action. 

As the pressure in the bellows decreases, 
spring tension acts to open the contacts. How- 
ever, since the force of the spring is opposed 
somewhat by the force of magnetic attraction, 
the contacts will not separate until a consider- 
able force is developed in the spring. This 
causes the contacts to snap open quickly so that 
arcing is again avoided. 

Toggle mechanisms are also frequently used 
as a means of causing the contacts to open and 
close with a snap action. Too, some controls 
employ a mercury switch as a means of over- 
coming the arcing problem. A typical mercury 
switch is illustrated in Fig. 13-10. As the glass 
tube is tilted to the right, the pool of mercury 
enclosed in the tube make contact between the 
two electrodes. As the bulb is tilted back to the 
left, contact is broken. The surface tension of 
the mercury provides the snap action necessary 
to prevent arcing. 

13-11. Space Control vs. Evaporator Con- 
trol. When the sensing element of the thermo- 
stat is located in the space or in the product, the 
thermostat controls the space temperature or 
product temperature directly. Likewise, when 
the sensing element is clamped to the evaporator, 
the thermostat controls the evaporator tempera- 
ture directly. In such cases, control of the space 
or product temperature is accomplished in- 



directly through evaporator temperature con- 
trol. Which of these two methods of control is 
the most suitable for any given application 
depends upon the requirements of the applica- 
tion itself. 

For applications where close control of the 
space or product temperature is desired, a 
thermostat which controls the space or product 
temperature directly will ordinarily give the best 
results. On the other hand, for applications 
where off-cycle defrosting is required and where 
minor fluctuations in the space or product tem- 
perature are not objectionable, indirect control 
of the space temperature by evaporator tempera- 
ture control is probably the better method. 

In order to assure complete defrosting of the 
evaporator, the evaporator must be allowed to 
warm up to a temperature of approximately 
37° or 38° F during each off cycle. When the 
thermostat controls the evaporator temperature, 
the cut-in temperature of the thermostat can be 
set for 38° F. Since the evaporator must warm 
up to this temperature before the compressor 
can be cycled on, complete defrosting of the 
evaporator during each off cycle is almost 
certain. On the other hand, when the thermo- 
stat controls the space temperature, there is no 
assurance that the evaporator will always warm 
up sufficiently during the off cycle to permit 
adequate defrosting. 

If we assume that the thermostat is properly 
adjusted, if the load on the system is relatively 
constant and the capacity of the system is 
sufficient to handle the load, no defrosting prob- 
lems are likely to arise with either method of 
temperature control. However, if the system 
is subject to considerable changes in load, 
defrosting problems are sometimes experienced 
in applications where the thermostat controls 
the space temperature. When the system is 



Glass tube 




Pool of 
mercury 



Contacts 
Fig. 13-10. Mercury contacts. 



236 PRINCIPLES OF REFRIGERATION 



operating tinder peak load conditions, the tem- 
perature of the space tends to remain above the 
cut-out temperature of the thermostat for 
extended periods so that running cycles are long 
and frost accumulation on the evaporator is 
heavy. Too, under heavy load conditions, the 
space temperature warms up to the cut-in 
temperature of the thermostat very quickly 
during the off cycle so that the off cycles are 
usually short. Frequently, the off cycles are too 
short to allow adequate defrosting of the eva- 
porator. In such cases, when the compressor 
cycles on again, the partially melted frost is 
caught on the evaporator and frozen into ice. 
Eventually the evaporator will be completely 
frozen over with ice, air flow over the evaporator 
will be severely restricted, and the system will 
become inoperative. 

13-12. Pressure Actuated Cycling Controls. 
Pressure actuated cycling controls are of two 
types: (1) low-pressure actuated and (2) high- 
pressure actuated. Low-pressure controls are 
connected to the low-pressure side of the system 
(usually at the compressor suction) and are 
actuated by the low-side pressure. High- 
pressure controls, on the other hand, are con- 
nected to the high-pressure side of the system 
(usually at the compressor discharge) and are 
actuated by the high-side pressure. 

The design of both the low-pressure and the 
high-pressure controls is similar to that of the 
remote-bulb thermostat. The principal differ- 
ence between remote-bulb thermostat and the 
pressure controls is the source of the pressure 
which actuates the bellows or diaphragm. 
Whereas the pressure actuating the bellows of 
the thermostat is the pressure of the fluid con- 
fined in the bulb, the pressures actuating the 
bellows of the low-and-high-pressure controls 
are the suction and discharge pressures of the 
compressor, respectively. Like the thermostat, 
both controls have cut-in and cut-out points 
which are usually adjustable in the field. 
13-13. High-Pressure Controls. High- 
pressure controls are used only as safety controls. 
Connected to the discharge of the compressor, 
the purpose of the high-pressure control is to 
cycle the compressor off in the event that the 
pressure on the high-pressure side of the system 
becomes excessive. This is done in order to 
prevent possible damaging of the equipment. 
When the pressure on the high-pressure side of 



the system rises above a certain, predetermined 
pressure, the high-pressure control acts to break 
the circuit and stop the compressor. When the 
pressure on the high-pressure side of the system 
returns to normal, the high-pressure control 
acts to close the circuit and start the compressor. 
However, some high-pressure controls are 
equipped with "lock-out" devices which require 
that the control be reset manually before the 
compressor can be started again. Although 
high-pressure controls are desirable on all 
systems, because of the possibility of a water 
supply failure, they are essential on systems 
utilizing water-cooled condensers. 

Since the condensing pressures of the various 
refrigerants are different, the cut-out and cut-in 
settings of the high-pressure control depend on 
the refrigerant used. 

13-14. Low-Pressure Controls. Low-pressure 
controls are used both as safety controls and as 
temperature controls. When used as a safety 
control, the low-pressure control acts to break 
the circuit and stop the compressor when the 
low-side pressure becomes excessively low and 
to close the circuit and start the compressor 
when the low-side pressure returns to normal. 
Like high-pressure controls, some low-pressure 
controls are equipped with a lock-out device 
which must be manually reset before the com- 
pressor can be started. 

Low-pressure controls are frequently used as 
temperature controls in commercial refrigeration 
applications. Since the pressure at the suction 
inlet of the compressor is governed by the 
saturation temperature of the refrigerant in the 
evaporator, changes in evaporator temperature 
are reflected by changes in the suction pressure. 
Therefore, a cycling control actuated by changes 
in the suction pressure can be utilized to control 
space temperature indirectly by controlling the 
evaporator temperature in the same way that 
the remote-bulb thermostat is used for this 
purpose. In such cases, the cut-in and cut-out 
pressures of the low-pressure control are the 
saturation pressures corresponding to the cut-in 
and cut-out temperatures of a remote-bulb 
thermostat employed in the same application. 
For example, assume that for a certain appli- 
cation the cut-in and cut-out temperature 
settings for a remote-bulb thermostat are 38° F 
and 20° F, respectively. If a low-pressure con- 
trol is used in place of the thermostat, the cut-in 






SYSTEM EQUILIBRIUM AND CYCLING CONTHOLS 237 



pressure setting for the lew-pressure control wj|] 
be SQpsia (the saturation pressure or R-12 
corresponding to a temperature of 38 V} and 
the cut-out pressure selling will be 36 psia (the 
saturation pressure of R-12 corresponding to a 
temperature or 20= F)-* 

As thie evaporator warms up during the off 
cycle, the pressure in the evaporator increases 
accordingly. When the pressure in the evapora- 
tor rises to the cut-in pressure setting of the low- 
pressure control, the low-pressure control acts 
to close the circuit and start the compressor. 
Very soon after the compressor start*. the 
temperature and pressure of ihe evaporator are 
reduced to approximately the design evaporator 
temperature and pressure and they remain nt 
this condition throughout most of the running 
cycle (see Fig. 13-8}, Near the end or the 
running cycle the evaporator temperature and 
pressure are gradually reduced below the design 
conditions. When the evaporator pressure is 
reduced to the cut-out pressure setting of the 
low-pressure control, the control Ads to break 
the circuit and stop the compressor. 

Since the refrigerant vapor undergoes ■ drop 
in pressure while flowing through the suction 
line, the pressure of the vapor at the suction 
inlet of the compressor is usually 2 or 3 lb less 
than the evaporator pressure. This is particu- 
larly true when the compressor is located 
some distance from the evaporator. Since ihe 
low-pressure control is actuated by the pressure 
at the suction inlet of the compressor, the 
pressure drop accruing in the suction line must 
be taken into account when the pressure control 
Bettings are made. To compensate for Ihe 
pressure loss in the suction line, the cut-out 
pressure setting is lowered by an amount equal 
to the pressure loss in the lino. For example, 
assuming a 3-1 b pressure loss in the suction line, 
when the pressure in the evaporator is 36 psia, 
the pressure at the suction inlet of the com- 
pressor will be 33 psia. Hence, if it is desired to 
cycle the compressor off when the pressure in 
the evaporator is reduced to 36 psia, the cut-out 
pressure of the low-pressure control is set for 
33 psia. In this instance, failure to make an 
nltowuiicc for the pressure loss in the Miction 

* When refrigerants other than R-12 are raed ui 
the system, the pressure setting! will be the m nar- 
ration pr aam e s of those refrigerants corresponding 
to the desired cut-in and cut-out temperature* 



line would cause the control to cycle the com- 
pressor when the pressure in the evaporator was 
reduced to only 39 psia rather than the desired 
36 psia. The system would have a tendency to 
short cycle because the differential is too small 
and unsatisfactory operation would result. 

Pressure loss in the suction line in no way 
ajfects the cut-in selling of the control. Since 
pressure drop is a function of velocity or how, 
there is no pressure drop in the suction line 
when the system is idle. As soon as the com- 
pressor cycles off, the pressure at the suction of 
compressor rises to the evaporator pressure so 
that at the time the compressor cycles on the 
pressure at the compressor inlet is the same 
as the evaporator pressure. Hence, the cut-in 
pressure setting of the control is made without 
regard for the pressure drop in the suction 
line. 

Since the low-pressure control controls eva- 
porator temperature rather than space tempera- 
ture, it is an ideal temperature control for 
applications requiring off-cycle defrosting. This 
is particularly true for '"remote" installations 
where the compressor is located some distance 
from the evaporator, tn such installations, low- 
pressure temperature control has a distinct 
advantage over thermostatic temperature con- 
trol in that it ordinarily results in a considerable 
saving in electrical wiring. Because of the 
remote bulb, the thermostat must always be 
located near the evaporator or space whose 
temperature is being controlled. This requires 
that a pair of electrical conductors be installed 
between the fk)it are and the condensing unit. On 
the other hand, the low-pressure control is 
located at the compressor near the power source 
so that the amount of control wiring needed is 
much less. 

13-15. Dual-Praiiur* Controls, A dual-pres- 
sure control Is a combination or the low-and- 
high- pressure controls in a single control. 
Ordinarily, only one set of electrical contact 
points are used in the control, although a 
separate bellows assembly is employed for each 
of the two pressures. A typical dual pressure 
control is shown in Fig. 13-11. This type of 
pressure control is frequently supplied as Stan* 
dard equipment on condensing units. 
13*16. The Pump-Down Cycle, A commonly 
used method of cycling the condensing unit, 
known as a "pump-down cycle," employs both a 









33S PRINCIPLES OF REFRIGERATION 




Rg. 11-11. Du*l pnmirt control, {Covrtmj 9< P«nn Contrail. 



Inc.) 



thermostat and a low-pressure control. In 1 
pump-dawn cycle, the space or evaporator tccrh 
perature is controlled directly by the thermostat. 
However, instead of starting and stopping the 
compressor driver, the thermostat nets 10 open 
and clou a solenoid valve installed in the liquid 
line, usually near the refrigerant flow control 
(Fig. 13-12). As the space or evaporator tem- 
perature is reduced to the cut-out temperature 
of the thermostat, the thermostat breaks the 
solenoid circuit, thereby de-energizing the 
solenoid and interrupting the Gow of liquid 
refrigerant to the evaporator- Continued 
operation of the compressor causes evacuation 
of the refrigerant from that portion of the 
system beyond the point where the refrigerant 
flow is interrupted by the solenoid. When the 
pressure in the evacuated portion of the system 
is reduced to [he cut-out pressure of the 
low pressure control, the low-pressure control 
breaks the compressor driver circuit and stops 



the compressor. When the temperature of the 
space or evaporator rises to tile cut-in tem- 
perature of the thermostat, the thermostat 
closes the solenoid circuit and energizes the 
solenoid, thereby opening the liquid line and 
permitting liquid refrigerant to enter the 
evaporator. Since the c^ipumtar is warm, 
the liquid entering the ■.-..ipuraujr vaporises 
rapidly so that the evaporator pressure rises 
immediately to the cut-in pressure of the 
low-pressure control, whereupon the low- 
pressure control closes the compressor driver 
circuit and start* the compressor. 

The advantages of the purip-down cycle are 
many. One of the most important ones being 
that the amount of refrigerant absorbed by the 
oil in the cranfceasc of the compressor during the 
olf cycle is substantially reduced. The problem 
of crankcase oil dilution by refrigerant absorp- 
tion during the off cycle is fully discussed in 
Chapter IS. 



STSTEM EQUILIBRIUM AND CYCLING CONTROLS 239 



13-17. Variation i in System Capacity. Si is 
worth while lo notice that balh the operating 
conditions and the capacity ol" a refrigerating 
system change as the load on I lie system changes. 
When the load on the system is heavy and I he 
space temperature is high, I he evaporator TD 
will be somewhat larger than the design evapo- 
rator TD and the capacity of the evaporator will 
be greater than the design evaporator capacity. 
Because of the higher evaporator capacity, the 
suction temperature will also be higher so that 
equilibrium is maintained between the vaporiz- 
ing and condensing sections of the system. 
Hence, under heavy load conditions, the system 
operating conditions arc Somewhat higher than 
the average design conditions and the system 
capacity is somewhat greater than the average 
design capacity. Obviously, the horsepower 
requirements of the compressor are greatest 
at this peak load condition and the compressor 
driver must be selected lo have sufficient horse- 
power lo meet these requirements. 

Conversely, when the load on the system is 
light, the space temperature wilt be lower (Kan 
the average design space temperature, the 
evaporator TD will be less than the design TD, 
and the suction temperature will be lower than 
the design suction temperature. Therefore, the 
system operaiing conditions wilt be somewhat 
tower than the average design operating con* 
ditiofls and the system capacity will be some- 
what less than the average design capacity 

During each running cycle the system passes 
through a complete series of operating condi- 
tions and capacities, the operaiing conditions 
and capacity being highest at the beginning of 
the running cycle when the space temperature 
is highest, and lowest at the end of the running 
cycle when the space temperature Is lowest. 
However, during most of the running cycle, a 



well-designed system will operate very nearly 
at the design conditions, 

13-10, Capacity Control. The importance of 
balancing the system capacity with the system 
load cannot be overemphasized. Any time the 
System capacity deviates considerably from the 
system load, unsatisfactory operating conditions 
will result. It has already been pointed out that 
good practice requires that the system be 
designed to have n capacity equal to or slightly 
in excess of the average maximum sustained 
load. This is done so that the system will have 
sufficient capacity to maintain the temperature 
and humidity at the desired level during 
periods of peak loading. Obviously, as the 
cooling load decreases, there is a lendency for 
the system to become oversized in relation to 
the load. 

In applications where the changes in the 
average system load are not great, capacity 
control is adequately accomplished by cycling 
the system on and oft as described in the 
preceding sections. In such cases, assuming 
that the cycling controls are property adjusted, 
the relative length of the on and off cycles wilt 
vary with the load on the system. During 
periods when the load is heavy, on cycles will 
be long and off cycles will be short, whereas 
during periods when the load is light, on cycles 
will be short and off cycles will be long. Natur- 
ally, the degree of variation in the length of 
the on and off cycles will depend on the degree 
of load fluctuation. However, since the system 
must always be designed to have sufficient 
capacity to handle the maximum load, when 
the changes in the system load are substantial, 
it is evident that the system will be considerably 
oversized when the load is at a minimum. A 
system which is oversized for the load will 
usually prove lo be as unsatisfactory as one that 



!,„ 



flew antral 



c 



Flj. I S>-|1. Pump-down cycle. 



I EvwofHw 



c 



hmmh i. :i i. 



VPii 



V 



ftiwef JinT" 



nwnmiii 




140 PRINCIPLES OF MFMGERATtON 




Fif, I J- 13. Evaporator iplit Into two lefmcnti lor 
capacity control. Clining tht wl«noid rtlv» In (hi 
llqwW tto* Of iff rtwfii A rtndtrt ihlt portion of 
th* •viporvcor irwpemiva. Th* cipicn, reduction 
obulnad U proportion*) to th* iiiriac* iru cydwl 



is undersized for the load. When the system is 
undersized for the load, the running: »"« will 
be excessive, the space temperature will be high 
Tor emended periods, and the off cycles will be 
too short to permit adequate dcfrosiing of ihe 
evaporator. On the other hand, where the 
system is oversized for the load, the off cycles 
will be loo long and the equipment running 
time wtll he insufficient to remove the required 
amount of moisture from the space. This will 
result in unsatisfactory {higher than normal) 
humidity conditions in the refrigerated space. 

For this reason, when changes in the system 
toad arc substantial, it is usually necessary to 
provide some means of automatically (or 
manually) varying the capacity of the system 
other than by cycling Ihe system on and off. 
This b true also of large installations when the 
size of the equipment renders cycling the system 
on and off impractical. 

There are many methods of bringing the 
refrigerating capacity Into balance with the 
refrigerating load. Naturally, the most suitable 
method in any one case will depend upon the 
conditions and requirements of the installation 
itself. Some installations require only one or 
two step* of capacity control, whereas others 
require a number of steps. Frequently, several 
methods are employed simultaneously in order 
to obtain the desired flexibility. Too, in some 
cases, it is necessary to impose an artificial 
load on the equipment to achieve the proper 
balance between the sensible (temperature 



reduction) and latent (nvuistun reduction) 
loads.* In applications where the latent load 
b too large a percentage of the total load, 
satisfaction of the latent load will result in 
ovcrcooling of the space utile** sensible heat is 
artificially introduced into ihe space. In such 
cases, the sensible heat is usually added to the 
space in the form of relic, u The air is first 
paased across a cooling coil and cooled to the 
temperature necessary to reduce the moisture 
content to the desired level ,md the air is then 
reheated to ihe required dr\ hulb temperature. 
The reheating is accomplished with steam or 
hot water coils, with electric Mrip-heatcrs, or 
with hot gas from the compressor discharge, 

In some installations, iIk- refrigerating capa- 
city of the system is adequately controlled by 
controlling ihe capacity of the compressor only. 
Since the flow rate of the refrigerant must be 
the same in all component*, any change in the 
capacity of any one component will automati- 
cally result in a similar .iiljustment in the 
capaci ly of al I the o ther com portents. Therefore, 
increasing or decreasing the capacity of the 
compressor will, in effect, increase or decrease 
the capacity of the entire sisicm, However, 
it is important to notice that with this method 
of capacity control the operating conditions of 
the system will change as the capacity of the 
system changes. 

Where it is desired to vnr\ ihe Capacity of the 
System without allowing the operating condi- 
tions of the system to change it is necessary lo 




Fig- 11-14. Cell drtultad (or lac* control. (Court**? 
Kifimrd Diviiion. Am»He»n Air Filter Company, 
Inc.) 

* This problem is usually rr,- - < < .it utc in the winter' 
lime when the transmission i» .11 gAin) load is light. 



SYSTEM EQUILIBRIUM AND CYCLING CONTROLS 341 



control both the evaporator capacity and the 
compressor capacity directly.* 

Some of the more common methods of con- 
trolling evaporator and compressor capacities 
are considered in the following sections. 
13-1?. Evaporator Capacity Control. Prob- 
ably the most effective method of providing 
evaporator capacity control it lo divide the 
evaporator into several separate sections or 
circuits which are individually controlled so that 
one or more sections or circuit* an be cycled 
out as the land decreases (Fig. 13-13}. Using 
this method, any percentage of the evaporator 
capacity con be cycled out in any desired 
number of sieps. The number and size of 
the individual evaporator sections depend on 
the number of steps of capacity desired and the 
percent change in capacity required per step, 
respectively. The arrangement of the evaporator 
sections or circuiting depends on the relation* 
ship of the sensible load to ihe total load at 11k 
various load conditions- Basically, two circuit 
arrangements are possible- Evaporator circuit- 
ing can be arranged lo provide cither "face" 
control of "depth" control, or both (Fig. 13-14 
and 11-15). When "face" control is used, the 
"sensible heat ratio" of the evaporator is not 




Hb> I J- IS, Coll circuit Ml for dapih control. (Cour* 
teijr Ksrmind DMtl&n, American Air Filter Company, 
Inc.). 

* Any reduction in ivucjii loud unj/tit lyalcni 
capacity will also have some elf«t on the capacity 
of the condenser and on the sin of the refrigerant 
lines. These topics lie d taunted in Chapter* J 4 
and 17, respectively. 




Fact damper 



Fig. It 14, Evaporator equipped with * (ace damper 
co vary the quantity otalrpasjin g ever chc evaporator. 
Al the dam par movn toward the cloud paiition. the 
rwiiunea aialrut which the Mower muat wsrk is 
Increased so that the total quantity of air circulated 



ofTected.t On the other hand, "depth" control 
always changes the sensible heat ratio. As a 
general rule, the more depth the evaporator has 
the greater is its Jatcnt cooling (moisture 
removal) capacity. Hence, as one or more rows 
of the evaporator are cycled out, the sensible 
heat ratio increases. 

Another common method of varying the 
evaporator capacity is to vary the amount of 
air circulated over (he evaporator through the 
use of "face" or "face-ond-by-pass" dampers. 
(Fig*. 1 3- 16 and 13-17) Muflispeed blowers can 
also be used for this purpose. Too, is some in- 
siances, mul lis peed blowers and dampers are u%d 
together if) order to provide the desired balance. 

In nearly all cases, application of any of Ihe 
foregoing methods of evaporator capacity 
control will necessitate simultaneous control of 
compressor capacity. 

13-M, Compressor Capacity Control. There 
in a number of different methods of controlling 
the capacity of reciprocating compressors. One 
method, already mentioned, is to vary the speed 
of the compressor by varying the speed of the 
compressor driver. When an engine or turbine 
a employed to drive the compressor. Ihe 
compressor capacity can he modulated over a 
wide range by governor control of the com- 
pressor driver. 

When an electric motor drives the compressor. 
Only two speeds are usually available SO that the 
compressor operates either at full capacity or at 
30% capacity. When more than two speeds ore 
desired, it is necessary to use two separate 
windings in the motor, in which case four speeds 
will be possible. 

; Ratio of the acntlbic coding capacity to the 
total cooling capacity. 



MI PWKCJPLES OF KEFRlGc RATION 



Mt 



i 



By- pass damp** 



! - 



I - 



\ 



Fk» dime** 

Pit- fMT. Evaporator aqulpped with fit* and by- 
pus dimptn lor dpicicy control. D*mp*r* if* 
JrttereoAnectad w tint by-pui damper opum wider 
a* ft« d»mp*r It tlowd eft. With thii iirinjerrmnt 
tht quantity of air puling e*»r til* traporstor can 
be rtfulated by allowing mart or leu »lr to by-pua 
iht avipcrator . Hcwtver, ragirdilMJ of tba petition 
of thi dim para, tht total quaintly of air circulated 
remain* practically tht aim*. 



Capacity control of multicylinder corn- 
pretwrt it frequently obtained by "unloading" 
one or more cylinders » that they become 
ineffective. One method of accomplishing this 
is to by-pas the discharge from one or more 
cylinder] buck into the suction tine as shown 
in Fig. 13-18. When the suction pressure drop* 
to a certain predetermined value a solenoid 
valve in the by-pass line, actuated by a pressure 
control, opens and alky** ihe discharge from 
one or more cylinders to flow through the 
by-put tine back into the suction line where it 
mixes with the incoming suction vapor. At 
long as (he suction pressure remains below the 
cut-in setting of the conirol. the discharge from 
the unloaded cylinders continues to by-pus to 
the suction line. When Ihe suction pressure 
rises to the cut-out setting of the pressure- 
con trot, the solenoid valve ,% dc-energixed and 
the by-pass tine is closed so that the compressor 
it returned to full capacity operation. 

Another method of unloading compressor 
cylinders is to depress the suction valves of the 
cylinder or cylinders to be unloaded so that they 
remain open during the compression (up) 
stroke. With the auction v.thes held open, die 
suction vapor drawn into the cylinder during 
the suction stroke is returned to the suction line 
during the compnuvDn stroke. A typical 
unloadcr of this type it shown in Fig. 13-19, 

The operation of the unloadcr mechanism is 
as follows: when the suction pressure falls (o 



the cut-in pressure of the prcuure control, the 
control energizes the solenoid valve and admits 
high-pressure gas from the condenser to the 
unloadcr piston which act* to depress ihe 
suction valves and bold them open. When 
the suction pressure rises to the cut-out pressure 
of the pressure control, the solenoid valve is de- 
energized and the unloadcr piston is relumed to 
the normal position. 

In addition to providing; opacity control, 
cylinder unloadcr* of all t;,pes art used to 
unload the compressor cylinders during com- 
pressor start-up so that the compressor starts 
in an unloaded condition, thereby reducing (he 
inrush current demand. 

When any of the capacity control methods 
described thus far are used. Ihe horsepower 
requirements of the compressor decrease as the 
capacity decreases, although not in the same 
proportions. 

Another method of controlling compressor 
capacity is to throttle the cm if pressor suction. 
However, since it reduces compressor capacity 
uiihout reducing compressor horsepower, this 
method it seldom used. 

Still another meant of con trolling compressor 
capacity which it employed with good results is 
to operate two or more compressors in parallel 
(Fig. 13-20). Individual low.. pressure controls 
ore used to cycle the compressor*. The cut-in 
and cut-out pressures of the mili vidua] controls 
are so adjusted that the corn pressors cycle off 
in sequence as the suction pressure decreases 
and cycle on in sequence when the suction 
pressure rises. Very often these compressors 
are equipped wilh cylinder unio;iders to provide 
additional steps of control Multiple com- 
pressor systems are discussed in detail in 
Chapter 20. 




He. I ML Schauta** diagram of cyilAoV by-pan. 



SYSTEM EQUILIBRIUM AND CYCUNG CONTROLS 243 



Typical 

solenoid valve 

(shown energized) 



Connect to 
discharge side 
of compressor 



Fig. 13-19. Condenser pres- 
sure actuated cylinder un- 
loader mechanism. (Courtesy 
Dunham-Bush, Inc.) 




13-21. Multiple-System Capacity Control. 

Another method of controlling capacity is to 
employ two or more separate systems. The 
evaporators for the separate systems may be in 
the same housing and air stream or they may 
be in separate housings and air streams. In 
either case, separate compressors and condensers 
are used, although in some instances the con- 
densers may be in a common housing. 




Fig. 13-20. Two compressors installed in parallel as 
a means of controlling compressor capacity. As the 
load diminishes, one compressor is cycled out to 
reduce the compressor capacity. Often, one com- 
pressor is equipped with cylinder unloaders to pro- 
vide additional steps of control. 



Valve plate 
Cylinder 



This method of capacity control is well 
suited to installations where only two steps of 
capacity control are required, as in chilling or 
combination chilling and storage applications. 
The use of two or more separate systems has the 
added advantage of providing a certain amount 
of insurance against losses accruing from 
equipment failure. Should one system become 
inoperative, the other can usually hold the load 
until repairs can be made. 

PROBLEMS 

1. Assuming a 2° F loss in saturated suction 
temperature due to refrigerant pressure drop 
in the suction line. 

(1) Select an air-cooled condensing unit to 
operate in conjunction with the natural 
convection evaporator in Problem 12-1. 

(2) Plot the evaporator and condensing unit 
capacities on a common graph and 
determine: 

(a) The saturated suction temperature at 
the point of system balance 

(b) The capacity of the system in Btu/hr at 
the point of system balance. 



14 

Condensers and 
Cooling Towers 



14-1. Condensers. Like the evaporator, the 
condenser is a heat transfer surface. Heat from 
the hot refrigerant vapor passes through the 
wails of the condenser to the condensing 
medium. As the result of losing heat to the 
condensing medium, the refrigerant vapor is 
first cooled to saturation and then condensed 
into the liquid state. 

Although brine or direct expansion refrig- 
erants are sometimes used as condensing 
mediums in low temperature applications, in 
the great majority of cases the condensing 
medium employed is either air or water, or a 
combination of both. 

Condensers are of three general types: 
(1) air-cooled, (2) water-cooled, and (3) evapo- 
rative. Air-cooled condensers employ air as the 
condensing medium, whereas water-cooled 
condensers utilize water to condense the 
refrigerant. In both the air-cooled and water- 
cooled condensers, the heat given off by the 
condensing refrigerant increases the temperature 
of the air or water used as the condensing 
medium. 

Evaporative condensers employ both air and 
water. Although there is some increase in 
the temperature of the air passing through the 
condenser, the cooling of the refrigerant in the 
condenser results initially from the evaporation 
of the water from the surface of the condenser. 
The function of the air is to increase the rate of 
evaporation by carrying away the water vapor 
which results from the evaporating process. 
14-2. The Condenser Load. Since the heat 
given up by the refrigerant vapor to the con- 



densing medium includes both the heat absorbed 
in the evaporator and the heat of compres- 
sion, the heat load on the condenser always 
exceeds that on the evaporator by an amount 
equal to the heat of compression. Since the 
work (heat) of compression per unit of refrig- 
erating capacity depends upon the compression 
ratio, the heat load on the condenser per unit of 
evaporator load varies with the operating 
conditions of the system. 

The quantity of heat liberated at the con- 
denser per minute per ton of evaporator 
capacity at various suction and condensing 
temperatures can be estimated from Charts 14-1, . 
14-2, and 14-3. Chart 14-1 applies to R-12 
systems, whereas Charts 14-2 and 14-3 apply to 
R-22 and R-717 (ammonia) systems, respec- 
tively. The values given are based on a simple 
saturated cycle. 

Example 14-1. An R-12 system, operating 
at a 15° F suction temperature, has a condens- 
ing temperature of 100° F. Determine the load 
on the condenser in Btu per minute per ton. 

Solution. In Chart 14-1, locate the 15° F 
suction temperature line at the base of the graph. 
Follow the line until it intersects the 100° F 
condensing temperature curve. The value on 
the left-hand index corresponding to this point 
is approximately 245 Btu per minute per ton. 

It is evident that for any given set of operating 
conditions there is a fixed relationship between 
the condenser load and the evaporator load. 
For instance, for the R-12 system described in 
Example 14-1, the relationship between the 
condenser load and the evaporator load is such 
that 245 Btu are liberated at the condenser for 
each 200 Btu taken in at the evaporator. 

Once the relationship between the condenser 
load and the evaporator load has been estab- 
lished for any given set of operating conditions, 
the total condenser load corresponding to any 
given total evaporator load can be easily 
computed. The following equation may be used : 



ax* 

He 



(14-1) 



where Q c = the condenser load in Btu/hr 
Q e = the evaporator load in Btu/hr 
q e = the condenser load in Btu/min/ton 

(from Fig. 14-15) 
q e = the evaporator load in Btu/min/ton 

(always 200 Btu) 



244 



CONDENSERS AND COOLING TOWERS 245 



Note: Q t may also be in Btu/min or in tons, 
in which case Q e will be in Btu/min or in tons, 
respectively. 

Example 14-2. For the system described in 
Example 14-1, determine the load on the con- 
denser if the load on the evaporator is 35,000 
Btu/hr. 

Solution. Applying 
Equation 14-1, the load = 35,000 x 245/200 
on the condenser, Q c = 42,875 Btu/hr 

It is important to notice that any increase or 
decrease in the load on the evaporator (system) 
will result in a proportional increase or decrease 
in the load on the condenser. 
14-3. Condenser Capacity. Since heat trans- 
fer through the condenser walls is by conduction, 
condenser capacity is a function of the funda- 
mental heat transfer equation: 



Q = A x U x D 



(14-2) 



where Q = the condenser capacity (Btu/hr) 

A = the surface area of the condenser 

(sqft) 
U = the transfer coefficient of the con- 
denser walls (Btu/hr/sq ft/° F) 
D — the log mean temperature difference 
between the condensing refrigerant 
and the condensing medium 
Examination of the factors in Equation 14-2 
will show that for any fixed value of U 
the capacity of the condenser depends on the 
surface area of the condenser and on the 
temperature difference between the condensing 
refrigerant and the condensing medium. It is 
evident also that for any one condenser of 
specific size and design, wherein the surface 
area and the U factor are both fixed at the time 
of manufacture, the capacity of the condenser 
depends only on the temperature differential 
between the refrigerant and the condensing 
medium. Therefore, for any one specific 
condenser, the capacity of the condenser is 
increased or decreased only by increasing or 
decreasing the temperature differential. Further- 
more, if it is assumed that the average tempera- 
ture of the condensing medium is constant, it 
follows that an increase or a decrease in the 
capacity of the condenser is brought about only 
by an increase or a decrease, respectively, in the 
condensing temperature. 



Since the condenser capacity must always be 
equal to the condenser load, it is evident from 
the foregoing that, for any given condensing 
load, the larger the surface area of the condenser, 
the smaller will be the required temperature 
differential and the lower will be the condensing 
temperature. Too, since the load on the 
condenser is always proportional to the load on 
the evaporator (system), any increase or decrease 
in the load on the evaporator will be reflected by 
an increase or a decrease, respectively, in the 
condensing temperature. 
14-4. Quantity and Temperature Rise of 
Condensing Medium. In both the air-cooled 
and water-cooled condensers, all the heat given 
off by the condensing refrigerant increases the 
temperature of the condensing medium. There- 
fore, in accordance with Equation 2-8, the 
temperature rise experienced by the condensing 
medium in passing through the condenser is 
directly proportional to the condenser load and 
inversely proportional to the quantity and 
specific heat of the condensing medium, viz: 



(T a - T x ) = 



ft 



M x C 



(14-3) 



where T x = the temperature of the air or water 
entering the condenser (7",) 
T 2 = the temperature of the air or water 
leaving the condenser (T x ) 
(r 2 — Tj) = the temperature rise (AT) experi- 
enced in the condenser 
Q, = the load on the condenser in Btu 

per hour 
M = the weight of air or water circulated 
through the condenser in pounds 
per hour 
C = the specific heat at constant 
pressure of the air or water 
Assuming that C has a constant value, for 
any given condenser load (Q,), Equation 14-2 
contains only two variables, M and AT, the 
value of each being inversely proportional 
to the value of the other, viz: 



M = 
AT = 



Q. 



C x AT 
C xM 



(14-4) 
(14-5) 



Therefore, for any given condenser load, if 
the temperature rise of the condensing medium 



246 PRINCIPLES OF REFRIGERATION 

is known, the quantity of condensing medium 
circulated through the condenser in pounds per 
hour can be determined by applying Equation 
14-4. Likewise, if the quantity circulated is 
known, the temperature rise through the con- 
denser can be computed by applying Equation 
14-5. 

Average specific heat values for air and water 
are 0.24 Btu/lb and 1 Btu/lb, respectively. By 
substituting the appropriate value for C, 
Equations 14-4 and 14-5 can be written to apply 
specifically to either water or air, viz: 



(14-6) 

(14-7) 
(14-8) 

(14-9) 



Since general practice is to express air and 
water quantities in cubic feet per minute (cfm) 
and in gallons per minute (gpm), respectively, 
it is usually desirable to compute condensing 
medium quantities in these units rather than in 
pounds per hour. 

To convert pounds of water per hour into 
gallons per minute, divide by 60 min to reduce 
pounds per hour to pounds per minute, and 
then divide by 8.33 lb per gallon to convert 
pounds per minute to gallons per minute, viz: 



for water 


-£ 




»-% 


and for air 


M Q ' 


"' 0.24 x AT 




IT Qs 




*" 0.24 x M 



gpm 



M(lb/hr) 



60 min x 8.33 lb/gal 



If these conversion factors are incorporated 
into Equation 14-6, the water quantity can be 
computed directly in gpm. The following equa- 
tion results: 

Q. 
8pm = 60 x 8.33 x Ar 

or, combining constants (60 x 8.33 

G, 



gpm 



500 X AT 



500), 
(14-10) 



convert from pounds per minute to cubic feet 

per minute, viz: 

M (lb/hr) x t>(cu ft/lb) 

cfm = rz — : 

60 min 

Assuming the specific volume of the air to be 
the specific volume of standard air (13.34 cu ft/ 
lb), incorporation of these conversion factors 
into Equation 14-7 results in the following: 
Q, x 13.34 cu ft/lb 



To reduce pounds of air per hour to cubic 
feet per minute, divide pounds per hour by 60 
min to determine pounds per minute and then 
multiply by the specific volume of the air to 



cfm = 



0.24 x 60 x AT 



or, combining constants (13.34/0.24 x 60 = 
1/1.08), 

Qs (14-11) 



cfm = 



1.08.x AT 



Example 14-3. If the load on a water- 
cooled condenser is 150,000 Btu/hr and the 
temperature rise of the water in the condenser 
is 10° F. What is the quantity of water cir- 
culated in gpm ? 



Solution. Applying Equation 
14-10, the water quantity in gpm 



150,000 



500 x 10 

= 30 gpm 

Example 14-4. TheJoad on a water-cooled 
condenser is 90,000 Btu/hr. If the quantity of 
water circulated through the condenser is 15 
gallons per minute, determine the temperature 
rise of the water in the condenser. 



Solution. Rearranging and 
applying Equation 14-10, AT 



90,000 
~ 500 x 15 
= 12°F 



Example 14-5. Thirty-six gallons of water 
per minute are circulated through a water- 
cooled condenser. If the temperature rise of the 
water in the condenser is 12° F, compute the 
load on the condenser in Btu/hr. 

Solution. Rearrang- 
ing and applying Equa- = 500 x Ar x gpm 
tion 14-10, the load = 500 x 12 x 36 
on Q, the condenser, = 216,000 Btu/hr 

Example 14-6. The load on an air-cooled 
condenser is 121,500 Btu/hr. If the desired 
temperature rise of the air in the condenser is 
25° F, determine the air quantity in cfm which 
must be circulated over the condenser. 



Solution. Applying Equa- 
tion 14-11, the air quantity in 
cfm 



121,500 
1.08 x 25 
4500 cfm 



CONDENSERS AND COOLING TOWERS 247 



Fig. 14-1. Water temperature 
rise through condenser. 



92°-».Water out 




Example 14-7. Three thousand cfm of air 
are circulated over an air-cooled condenser. If 
the load on the condenser is 64,800 Btu/hr, 
compute the temperature rise of the air passing 
over the condenser. 



64,800 



Solution. Rearranging and 
applying Equation 14-1 1, AT = -^ -- ^ 

= 20°F 

For any given condenser and condenser 
loading, the condensing temperature of the 
refrigerant in the condenser will depend only 
upon the average temperature of the con- 
densing medium flowing through the condenser. 
The lower the average temperature of the con- 
densing medium the lower is the condensing 
temperature. For example, assume that the 
size and loading of a condenser are such that 
the required mean temperature differential 
between the refrigerant and the condensing 
medium is 15° F. If the average temperature of 
the condensing medium is 90° F, the condensing 
temperature will be 105° F (90 + 15), whereas if 
the average temperature of the condensing 
medium is 85° F, the condensing temperature 
will be 100° F (85 + 15). 

The average temperature of the condensing 
medium flowing through the condenser depends 
upon both the initial temperature of the 
condensing medium entering the condenser and 
the temperature rise experienced in the con- 
denser. Since the temperature rise of the 
condensing medium decreases as the flow rate 
increases, the greater the quantity of condensing 
medium circulated the lower is the average 
temperature of the condensing medium. There- 
fore, for any given condenser loading, the 



greater the flow rate of the condensing medium 
the lower will be the condensing temperature. 
14-5. Condenser Application. As a general 
rule, for any given condenser load, the size of 
the condenser and the quantity of condensing 
medium circulated will depend upon the 
entering temperature of the condensing medium 
and upon the desired condensing temperature. 
A careful analysis of the data in Sections 14-3 
and 14-4 will show that the condensing tempera- 
ture of the refrigerant in the condenser is a 
function of three variables: (1) the entering 
temperature of the condensing medium, (2) 
the temperature rise in the condenser, and (3) the 
temperature difference between the refrigerant 
and the condensing medium. This relationship 
is illustrated in Fig. 14-1. 

Recalling that the temperature rise in the 
condenser varies inversely with the flow rate of 
the condensing medium and that the tempera- 
ture differential between the refrigerant and the 
condensing medium varies inversely with the 
size (surface area) of the condenser,* it is 
evident that: 

1. For any given condensing surface and 
flow rate, die condensing temperature will 
increase or decrease as the entering temperature 
of the condensing medium increases or de- 
creases. 

2. For any given entering temperature, the 
larger the condensing surface and the higher the 
flow rate, the lower will be the condensing 
temperature. 

3. For any given entering temperature, the 

* Assuming the transfer coefficient to be con- 
stant. 



248 PRINCIPLES OF REFRIGERATION 



amount of condensing surface required for a 
given condensing temperature decreases as the 
flow rate of the condensing medium increases. 

With regard to the latter statement, this 
means in effect that the same condensing 
temperature can be maintained with either a 
small condensing surface and a high flow rate 
or a large condensing surface and a low flow 
rate. However, it should be recognized that the 
flow rate of the condensing medium is fixed 
within certain limits by the size and design of 
the condenser. If the flow rate through the 
condenser is too low, flow will be streamlined 
rather than turbulent and a low transfer coeffi- 
cient will result. On the other hand, if the flow 
rate is too high, the pressure drop through the 
condenser becomes excessive, with the result 
that the power required to circulate the con- 
densing medium also becomes excessive. 

Since the design entering temperature of the 
condensing medium is usually fixed by condi- 
tions beyond the control of the system designer, 
it follows that the size and design of the con- 
denser and the flow rate of toe condensing 
medium are determined almost entirely by the 
design condensing temperature. 

Although low condensing temperatures are 
desirable in that they result in high compressor 
efficiency and low horsepower requirements for 
the compressor, this does not necessarily mean 
that the use of a large condensing surface and a 
high flow rate in order to provide a low con- 
densing temperature will always result in the 
most practical and economical installation. 
Other factors which must be taken into account 
and which tend to limit the size of the condenser 
and/or the quantity of condensing medium 
circulated are initial cost, available space, and 
the power requirements of the fan, blower, or 
pump circulating the condensing medium. Too, 
where water is used as the condensing medium 
and the water leaving the condenser is wasted 
to the sewer (see Section 14-9), the availability 
and cost of the water must also be considered. 

The limitations imposed on condenser size by 
the factors of initial cost and available space 
are self-evident. As for the power requirements 
of the fan, blower, or pump circulating the 
condensing medium, it has already been stated 
that the power required to circulate the con- 
densing medium increases as the flow rate 



increases. If the flow rate is increased beyond a 
certain point, the increase in the power required 
to circulate the condensing medium will more 
than offset the reduction in the power require- 
ments of the compressor accruing from the 
increased flow rate. Therefore, the quantity of 
condensing medium which can be economically 
circulated is limited by the power requirements 
of the fan, blower, or pump. 

Obviously, the optimum flow rate for the 
condensing medium is the one which will result 
in the lowest over-all operating costs for the 
system. This will vary somewhat with the 
conditions of the individual installation, being 
influenced by the type of application, the size 
and type of condenser used, fouling rates, and 
the design conditions for the region, along with 
such practical considerations as the cost and 
availability of water, utility costs, local codes 
and restrictions, etc. For example, since good 
system efficiency prescribes lower condensing 
temperatures for low temperature applications 
than for high temperature applications, it 
follows that for the same condenser load the 
optimum condensing medium flow rate will 
usually be higher for a low temperature appli- 
cation than for a high temperature application. 
Too, where the entering temperature of the 
condensing medium is relatively high, larger 
condensing surfaces and higher flow rates 
are required to provide reasonable condensing 
temperatures than where the entering tempera- 
ture of the condensing medium is lower. 
14-6. Air-Cooled Condensers. The circula- 
tion of air over an air-cooled condenser may be 
either by natural convection or by action of a 
fan or blower. Where air circulation is by 
natural convection, the air quantity circulated 
over the condenser is low and a relatively large 
condensing surface is required. Because of 
their limited capacity natural convection con- 
densers are used only on small applications, 
principally domestic refrigerators and freezers. 

Natural convection condensers employed on 
domestic refrigerators are usually either plate 
surface or finned tubing. When finned tubing 
is used, the fins are widely spaced so that little 
or no resistance is offered to the free circulation 
of air. Too, wide fin spacing reduces the 
possibility of the condenser being fouled with 
dirt and lint. 

The plate-type condenser is mounted on the 



CONDENSERS AND COOLING TOWERS 249 



back of the refrigerator in such a way that an 
air flue is formed to increase air circulation. 
Finned tube condensers are mounted either on 
the back of the refrigerator or at an angle 
underneath the refrigerator. Regardless of 
condenser type or location, it is essential that the 
refrigerator be so located that air is permitted 
to circulate freely through the condenser at all 
times. Too, warm locations, such as one 
adjacent to an oven, should be avoided when- 
ever possible. 

A number of domestic freezer manufacturers 
utililize the outer shell of the freezer (outside 
wall surface) as a condensing surface. This is 
accomplished by bonding bare tubing to the 
inside surface of the outer shell so that the 
entire outer shell becomes a plate-type heat 
transfer surface. The use of these "wrap- 
around" condensers permits a considerable 
reduction in the size of the freezer (6 to 8 in. on 
both length and width), not only because it 
eliminates the space ordinarily required for the 
condenser but also because it allows the use of 
3 to 4 in. of insulation in the walls where 
normally 6 to 8 in. is required in order to prevent 
moisture from condensing on the outside surface 
of the freezer. The slightly higher operating 
costs which accrue as a result of reducing the 
amount of wall insulation is more than offset 
by the savings in space that this practice makes 
possible. 

Air-cooled condensers employing fans or 
blowers to provide "forced-air" circulation can 
be divided into two groups according to the 
location of the condenser: (1) chassis-mounted 
and (2) remote. 

A chassis-mounted air-cooled condenser is 
one that is mounted on a common chassis with 
the compressor and compressor driver so that 
it is an integral part of the air-cooled "con- 
densing unit" (Fig. 6-12). Although chassis- 
mounting of the air-cooled condenser makes 
possible a very compact, completely self- 
contained condensing unit which is ideally 
suited for use on small commercial fixtures, 
this arrangement has certain inherent disadvan- 
tages which make chassis-mounting impractical 
in larger applications. 

The principal disadvantage of chassis- 
mounted air-cooled condensers is that the 
physical size of the condenser is limited to 
the dimensions of the chassis. Because of the 



limitation in physical size, chassis-mounted 
condensers of the type shown in Fig. 6-12 are 
rarely found in capacities larger than 2 tons.* 

Another disadvantage of the chassis-mounted 
air-cooled condenser is their susceptibility to 
fouling. Since most condensing units are 
mounted on the floor, the condenser air tends 
to sweep across the floor so that dirt, lint, and 
other foreign materials are picked up from the 
floor and carried to the surface of the condenser, 
thereby "clogging" the condenser and restricting 
the air flow. 

Too, on "open-type" air-cooled condensing 
units the condenser fan is usually mounted on 
the shaft of the compressor driver (Fig. 6-12). 
Naturally this limits both the size and the 
location of the fan so that the quantity of air 
circulated over the condenser is always less 
than that which would produce maximum 
efficiency at full load conditions. Notice also 
that, because of the fan location, the distri- 
bution of the air over the condenser surface is 
very uneven, being much greater on the end of 
the condenser directly in front of the fan. 

Remote air-cooled condensers are used in 
sizes from 1 ton up to 100 tons or more and may 
be mounted either indoors or outdoors. When 
located indoors, provisions must be made for an 
adequate supply of outside air to the condenser 
(Fig. 14-2). If the condenser is installed in a 
warm location, such as in an attic or boiler 
room, ducts should be used to carry the air into 
the condenser and back to the outside. Because 
of the large quantity of air required, only the 
smaller sizes are mounted indoors. 

When located outdoors, the air-cooled con- 
denser may be mounted on the ground, on the 
roof, or on the side of a wall, with roof locations 
being the most popular. Typical wall and roof 
installations are shown in Figures 14-3 and 
14-4, respectively. In all cases, the condenser 
should be so oriented that the prevailing winds 
for the area in the summertime will aid rather 
than retard the action of the fan. In the event 
that such orientation is not possible, wind 
deflectors should be installed on the discharge 
side of the condenser (Fig. 14-5). 

* This is the approximate condenser capacity 
required on a 3 hp, commercial, air-cooled con- 
densing. Approximately 25 % of the motor horse- 
power is required to drive the fan. Naturally, this 
reduces the horsepower available to the compressor. 



250 PRINCIPLES OF REFRIGERATION 



Ceiling 



Purge valve 




Fig. 10-2. Indoor installation of 
air-cooled condenser. (Cour- 
tesy Kramer Trenton Com- 
pany.) 



/ Compressor'' 

* Locate receiver below unicon outlet 

One significant outgrowth of the remote air- 
cooled condenser has been the development of 
a new type of air-cooled condensing unit which 
is designed specifically for remote installation. 
The air-cooled condensing unit illustrated in 
Fig. 14-6 is typical of these newer designs. This 




type is rapidly gaining in popularity and is now 
available in almost any desired capacity. 
14-7. Air Quantity and Velocity. For an 
air-cooled condenser there is a definite relation- 
ship between the size (face area) of the con- 
denser and the quantity of air circulated in that 
the velocity of the air through the condenser is 
critical within certain limits. Good design pre- 
scribes the minimum air velocity that will pro- 
duce turbulent flow and a high transfer coefficient. 
Increasing the air velocity beyond this point 
causes an excessive pressure drop through the 
condenser and results in an unnecessary increase 
in the power requirements of the fan or blower 
circulating the air. 

The velocity of the air passing through an air- 
cooled condenser is a function of the free face 
area of the condenser and the quantity of air 
circulated. The relationship is given in the 
following equation: 



Air velocity (fpm) = 



Air quantity (cfm) 
Free face area (sq ft) 



Fig. 14-3. Remote air-cooled condensers installed on 
outside wall. (Courtesy Kramer Trenton Company.) 



The free face area of the condenser is the area 
of the free air spaces between the tubes and fins. 
The actual free area per unit of face area varies 
with the design of the condenser, being depend- 
ent upon the size, number, and arrangement of 
the tubes and fins. 



CONDENSERS AND COOLING TOWERS 25 1 




Fig. 14-4. Remote air-cooled condensers mounted on roof. (Courtesy Dunham-Bush, Inc.) 



Normally, air velocities over air-cooled con- 
densers are between 500 and 1000 fpm. How- 
ever, because of the many variables involved, 
the optimum air velocity for a given condenser 
design is best determined by experiment. For 
this reason, most air-cooled condensers come 
from the factory already equipped with fans or 
blowers so that the air quantity and velocity 
over the condenser are fixed by the manu- 
facturer. In all cases, to realize peak perform- 
ance from an air-cooled condenser, the 



manufacturer's recommendations as to the air 
quantities should be carefully followed. 
14-8. Rating and Selection of Air-Cooled 
Condensers. Capacity ratings for air-cooled 
condensers are usually given in Btu/hr for 
various operating conditions. It has already 
been shown that since the surface area and the 
value of U are fixed at the time of manufacture, 
the capacity of any one condenser depends only 
on the mean temperature difference between 
the air and the condensing refrigerant. Since 



Fig. 14-5. Remote air-cooled 
condensers equipped with 
wind deflectors. (Courtesy 
Kramer Trenton Company.) 




252 PRINCIPLES OF REFRIGERATION 




Fig. 14-6. Air-cooled condensing unit designed for remote installation. Notice generous size of condenser. 
(Courtesy Kramer Trenton Company.) 



most air-cooled condensers come equipped with 
fans or blowers, the quantity of air circulated 
over the condenser is also fixed so that the 
average temperature of the air passing over the 
condenser depends only on the dry bulb tem- 
perature of the entering air and the load on the 
condenser. Obviously, then, in such cases, the 
capacity of the condenser is directly proportional 
to the temperature difference between the dry 
bulb temperature of the entering air and the 
condensing temperature. This temperature 
differential is often referred to as the "tempera- 
ture split" in order to distinguish it from the 
mean effective temperature differential.* 

Table R-13 is a typical manufacturer's rating 
table for air-cooled condensers. The basic 
ratings given in Table R-13A are based on 90° F 
ambient air temperature, 120° F condensing 
temperature, and 40° F evaporating tempera- 
ture. For other design conditions multiply the 
basic rating from Table R-13 A by the correction 
factors for variation in evaporating temperatures 
(Table R-13B) and for variation in entering air 
and condensing temperatures (Table R-13Q. 

* The temperature split is always proportional to 
the METD. 



In order to select a condenser from the rating 
tables, the following design data must be known : 

(1) The design suction and condensing tem- 
peratures 

(2) The compressor capacity in Btu/hr 

(3) The design outdoor dry bulb temperature 
(use values in Table 10-6A. Round off to 
next highest multiple of 5) 

Example 14-8. From Table R-13, select 
an air-cooled condenser for a compressor 
having a capacity of 75,000 Btu/hr if the design 
evaporating and condensing temperatures are 
20° F and 110° F, respectively, and the outdoor 
design dry bulb for the region is 90° F. 

Solution. From Table 
R-l 3B, the correction fac- 
tor for 20° F suction tem- 
perature = 0.95 

From Table R-12C, the 
correction factor for con- 
densing temperature of 
110° F and entering air 
temperature of 90° F = 0.665 

Required capacity of 
condenser at basic rating 75 000 

conditions - 0.95 x 0.665 

= 11 8,700 Btu/hr 



CONDENSERS AND COOLING TOWERS 253 



From Table R-13A, select condenser Model 
#BD1000 which has a capacity of 120,000 
Btu/hr at the basic rating conditions. 

Experience has shown that as a general rule 
selecting an air-cooled condenser on the basis 
of a condensing temperature of 1 10° F will 
result in the most economical condenser size. 
Hence, the actual size of the condenser selected 
will depend upon the outdoor design dry bulb 
temperature for the region in question. The 
higher the dry bulb temperature, the larger the 
condenser required. For example, for a con- 
densing temperature of 1 10° F, if the dry bulb 
temperature is 85° F, the condenser can be 
selected for a 25° F temperature split, whereas 
if the dry bulb temperature is 90° F, the con- 
denser must be selected for a 20° F temperature 
split, which will require a larger size. 
14-9. Water-Cooled Condenser Systems. 
Systems employing water-cooled condensers can 
be divided into two general categories: (1) 
waste-water systems and (2) recirculated water 
systems. In waste-water systems the water 
supply for the condenser is usually taken from 
the city main and wasted to the sewer after 
passing through the condenser (Fig. 14-7). In 
recirculated water systems the water leaving the 
condenser is piped to a water cooling tower 
where its temperature is reduced to the entering 
temperature, after which the water is recircu- 
lated through the condenser (Fig. 14-8). 

Naturally, where the condenser water is 
wasted to the sewer, the availability and cost of 
the water are important factors in determining 
the quantity of water circulated per unit of con- 
denser load. As a general rule, an economical 
balance between water and power costs pre- 
scribes a water flow rate of approximately 1.5 
gal per minute per ton of capacity. 

The high cost of water, along with limited 
sewer facilities and recurring water shortages in 
many regions, has tended to limit waste-water 
systems to very small sizes. Too, many cities 
have placed severe restrictions on waste-water 
systems, particularly where the water supply is 
taken from the city main and wasted to the 
sewer. 

When the condenser water is recirculated the 
power required to circulate the water through 
the water system must be taken into account in 
determining the water flow rate. Experience has 



shown that, in general, a water flow rate of 
between 2.5 and 3 gal per minute per ton usually 
provides the most economical balance between 
compressor horsepower and pump horsepower. 

In some instances, the water supply for a 
waste-water system is taken from a well or from 
some nearby body of water, such as a river, 
lake, pond, etc., in which case both the cost of 
the water and the pumping horsepower must be 
considered in determining the optimum water 
flow rate. 

To a large extent, the quantity of water cir- 
culated through the condenser determines the 
design of the water circuit in the condenser. 
Since heat transfer is a function of time, it 
follows that where low water quantities necessi- 
tate a high temperature rise in the condenser, 
the water must remain in contact with the con- 
densing refrigerant for a longer period than 
when the water flow rate is high and the tem- 
perature rise required is smaller. Hence, where 
the water flow rate is low, the number of water 
circuits through the condenser are few and the 
circuits are long so that the water will remain in 
the condenser for enough time to permit the 
required amount of heat to be absorbed. On 
the other hand, when the flow rate is high and 
the temperature rise low, more circuits are used 
and the circuits are shorter in order to reduce 
the pressure drop to a minimum. This is illu- 
strated in Figs. 14-9a and 14-96. In Fig. 14-9a, 
the two water circuits through the condenser are 
connected in series for a low flow rate and a high 
temperature rise. The water enters through 
opening A and leaves through opening C. 
Opening B is capped. In Fig. 14-9ft, the two 
water circuits are connected in parallel for a high 
flow rate and a low temperature rise. The water 
enters through opening B and leaves through 
openings A and C. 

In designing the condenser water circuit 
particular attention must be given to the water 



Suction 



/Water regulating valve 
Warm mater out _^. 




Fig. 14-7. Waste water system. 



254 PRINCIPLES OF REFRIGERATION 




Hot gas in 



Fig. 14-8. Recirculating water 
system. 



Pump-^ 



velocity and pressure drop through the con- 
denser. In all cases the minimum permissible 
velocity is that which will produce turbulent 
flow and a high transfer coefficient. Since 
pressure drop is a function of velocity, the 
pressure drop through the condenser increases 
as the water velocity increases. For this reason, 
the maximum permissible velocity in any one 
case is usually determined by the allowable 
pressure drop. * For waste-water systems, where 




to) 



Water out 


//t\rtsJDhJjt\ 


C Water in 








| *■* >* ** '*' 




B 


~U A\ A\ A\ A\ 












\y \y \y_\y w 


A Water in 



<b) 

Fig. 14-9. (o) Water circuit connected for series 
flow, (b) Water circuit connected for parallel flow. 

* Excessive velocity will usually cause erosion of 
the water tubes, particularly at points where the 
water changes direction. The maximum velocity 
recommended by Air Conditioning and Refrigeration 
Institute (ART) is 8 fps. 



the water is forced through the condenser by 
city main pressure, the pressure drop through 
the condenser is not critical as long as it is 
within the limits of turbulent flow and the avail- 
able city main pressure. In such cases, high 
velocities are recommended in order to take 
advantage of the higher transfer coefficient. On 
the other hand, when the water is circulated by 
action of a pump, a high pressure drop through 
the condenser will increase the pumping head 
and the power required to circulate the water. 
Therefore, for recirculating water systems, the 
optimum water velocity is one which will pro- 
vide the most economical balance between a 
high transfer coefficient and a low pumping 
head. 

In Figs. 14-9a and 14-9A, it is of interest to 
notice that for the same flow rate the velocity 
and pressure drop through the circuit arrange- 
ment in Fig. 14-9a are approximately four times 
as great as that through the circuit arrangement 
in Fig. 14-96. Too, because of the higher velo- 
city, the transfer coefficient is somewhat higher 
for the condensing surface in Fig. 14-9a and less 
condensing surface is required for the same heat 
transfer capacity. 

14-10. Fouling Rates. Another factor which 
must be considered in selecting a water-cooled 
condenser is fouling of the tube surface on the 
water side. The fouling is caused primarily by 
mineral solids which precipitate out of the water 
and adhere to the tube surface. The scale thus 
formed on the tube not only reduces the water 
side transfer coefficient, but it also tends to 



CONDENSERS AND COOLING TOWERS 255 



restrict the water tube and reduce the quantity 
of water circulated, both of which will cause 
serious increases in the condensing pressure. 

In general, the rate of tube fouling is in- 
fluenced by: (1) the quality of the water used 
with regard to the amount of impurities con- 
tained therein, (2) the condensing temperature, 
and (3) the frequency of tube cleaning with 
relation to the total operating time. 

Most manufacturers of water-cooled con- 
densers give condenser ratings for clean tubes 
and for four stages of tube fouling in accordance 
with the scale factors given in Table 14-1 for 
various types of water. These scale factors are 
an index of the reduction in the tube transfer 
coefficient resulting from the scale deposit. In 
selecting a water-cooled condenser, a minimum 

Refrigerant 
" vapor in 



this arrangement, some air-cooling of the refrig- 
erant is provided in addition to the water- 
cooling. Counterflowing of the fluids in any 
type of heat exchanger is always desirable since 
it results in the greatest mean temperature 
difference between the fluids and, therefore, the 
highest rate of heat transfer. 

Several types of double-tube condensers are 
shown in Figs. 14-11 and 14-12. The type 
shown in Fig. 14-11 can be cleaned mechan- 
ically by removing the end-plates (inset). The 
type shown in Fig. 14-12 is cleaned by circu- 
lating approved chemicals through the water 
tubes (see Section 14-23). 

Equipped with water-regulating valves (Sec- 
tion 14-20), double-tube condensers make ex- 
cellent "booster" condensers for use with 






Water out 



Fig. 14-10. Double -tube 
water-cooled condenser. 




Water in^ 



P 



scale factor of 0.0005 should always be used. 
Under no circumstances should a condenser be 
selected on the basis of clean tubes. However, 
when the condensing temperature is low (leaving 
water temperature less than 100° F) and the 
condenser tubes are to be cleaned frequently, 
the fouling factor from Table 14-1 may be 
reduced to the next lowest value. The use of 
scale factors will be illustrated later in the 
chapter. 

14-11. Water-Cooled Condensers. Water- 
cooled condensers are of three basic types: 
(1) double-tube, (2) shell-and-coil, and (3) shell- 
and-tube. 

As its name implies, the double-tube con- 
denser consists of two tubes so arranged that 
one is inside of the other (Fig. 14-10). Water is 
piped through the inner tube while the refrig- 
erant flows in the opposite direction in the 
space between the inner and outer tubes. With 



Condensed 
refrigerant out 



chassis-mounted air-cooled condensers during 
periods of peak loading. Since the water valve 
can be adjusted to open and allow water to flow 
through the condenser only when the con- 
densing pressure rises to some predetermined 
level, the amount of water used is relatively 
small in comparison to the savings in power 
afforded by the increased compressor effi- 
ciency. 

The shell-and-coil condenser is made up of 
one or more bare-tube or finned-tube coils 
enclosed in a welded steel shell (Fig. 14-9). The 
condensing water circulates through the coils 
while the refrigerant is contained in the shell 
surrounding the coils. Hot refrigerant vapor 
enters at the top of the shell and condenses as it 
comes in contact with the water coils. The con- 
densed liquid drains off the coils into the bottom 
of the shell, which often serves also as the 
receiver tank. Care should be taken not to 



256 PRINCIPLES OF REFRIGERATION 





Fig. 14-11, Double-pipe condenser* wlih mechanically deanable tube*. {Courtwy Hilnead and Mitchell.) 



overcharge the system with refrigerant since an 
excessive accumulation of liquid in the ton- 
denser will tend to cover too much of the con- 
densing surface and cause an increase in Frit 
discharge temperature and pressure. 

Most sheli-and-coil condensers are equipped 
with a split water circuit. The two parts of the 
circuit are connected in series for waste-water 
systems (Fig. J 4-96) and in parallel for recircu- 
lating systems (Fig. i4-9a). As a general rule, 
shcll-and-coil condensers arc used only for small 
installations up to approximately 10 tons 
capacity, 

Shelt-and-coil condensers are cleaned by cir- 
culating an approved chemical through the 
water coils. 

The sbcll-and-tube condenser consists of a 
cylindrical steel shell in which a number of 
straight tubes are arranged in parallel and held 
in place at the ends by tube sheets. Construc- 
tion is almost identical to that of the flooded- 
type shell-and-tube liquid chiller. The con- 
densing water is circulated through the tubes, 
which may be either steel or copper, bare or 
extended surface. The refrigerant is contained 
in the steel shell between the tube sheets. Water 
circulates in the annular spaces between the tube 
sheets and the end-piates, the end-plates being 
baffled to act as manifolds to guide the water 
flow through the tubes. The arrangement of the 
end-plate bathing determines the number of 
passes the water makes through the condenser 



from one end to the other before leaving the 
condenser. The number of passes may be as few 
as two or as many as twenty. 

For any given total number of lubes, the 
number of tubes per pass varies inversely with 
the number of passes. For example, assuming 
that a condenser has a total of forty tubes, if 
there are two passes, the number of tubes per 
pass is twenty, whereas if there are four passes, 
the number of tubes per pass is ten. 

It is important to notice that for the same 
total number of tubes and the same water 
quantity, the velocity of the water and the 
pressure drop through the condenser will be four 
times as great for a four-pass condenser as for a 
two-pass condenser. Because of the higher velo- 
city the transfer coefficient will be higher for the 
four-pass condenser and a smaller condensing 
surface will be required for a given heat transfer 
capacity. However, on the other hand, because 
of the high pressure drop, the power required to 
circulate the water will be greater. Hence, for a 
waste- water system, the- four-pass condenser is 
probably the best selection, whereas for a 
recirculating system, the two-pass condenser is 
probably the better of the two.* 

Shcil-and-fube condensers are available in 
capacities ranging from 2 tons up to several 

* This example is intended only In illustrate the 
principles of design and should not be construed 
to mean that four-pass condensers are undesirable 
for recirculating systems. 



CONDENSERS AND COOLING TOWERS 157 





Hg. 14-12. Typical double-pipe condenser cortfigur«iofl*. (fl) Trombone configuration, (t) HeUlt configura- 
tion. (Courtesy Edwardi Engineering Corporation.) 



hundred tons or more. Shell diameters range 
from approximately 4 in, up to 60 in,, whereas 
tube length varies from approximately j ft to 20 
ft. The number and the diameter or the tubes 
depend on the diameter of the shell. Tube 
diameters of g in. through 2 in. are common, 
whereas the number of Lubes in the condenser 
varies from as few as six or eight to as many as a 
thousand or more. The end-plates of the con- 
denser are removable to permit mechanical 
cleaning of the water tubes. 

Single-pass, vertical sheH-and-tube condensers 
are sometimes employed on large ammonia 
installations. The construction of the vertical 
she! I -and- lube condenser is similar to that of the 
vertical shcll-and-tubc chiller illustrated in Fig. 
1 1 -44. The vertical condenser is equipped * i th a 
water box at the top to distribute the water to 
the tubes and a drain at the bottom to carry the 
water away. Bach tube is equipped at the top 
with a distributor fitting which imparts a rotating 
motion to the water to assure adequate wetting 
of the lube. The hot refrigerant vapor usually 
enters at the side of the shell near the middle of 
the condenser and the liquid leaves the con- 
denser at the side of the shell near the bottom. 
The height of vertical shell-and-tube condensers 
ranges from 12 ft to IE ft. The tubes are 
mechanically clcanabk. 

14-12- Rating and Selection of Water' 
Cooled Condensen.* The ratings shown in 

* The maLcriiil in this section is reprinted directly 
from the manufacturer's catalog, the only alter* 
ation being the. table designations. Courtesy of 
Acme Industries, Inc. 



Table R-I4 are based on condensing tempera- 
tures of 102° and 105^ F, 20" and 10" water rise 
and 0.0005 scale factor which is the minimum 
recommended in ARf standards. 

Where other conditions exist, the following 
procedure should be followed in selecting the 
proper condenser. 

Condensers must not be selected for less than 
0.5 gpm per tube below which streamline instead 
of turbulent water flow occurs. ART standards 
indicate that the water velocity should not 
exceed E fps which is 5.75 gpm per tube for 
Acme STF and SRF condensers. 

It is necessary to have the following informa- 
tion to select a proper condenser : 

1. Total tons (low side). 

2. Evaporator temperature. 

3. Condensing temperature. 

4. Water temperature "in." 

5. Water temperature "out," or gpm avail- 
able. 

6. Type of water or required scale factor. 

Then proceed as follows : 

1 . Determine the corrected tons to be used in 
selecting the proper condenser by reference to 
Fig. 2, Table R-I4. The factor obtained for the 
desired evaporator temperature and condensing 
temperature is multiplied by the actual tons to 
obtain corrected tons. 

2. Determine the water temperature rise and 
gpm per ton. Knowing cither factor, the other 
may be obtained by reference to Fig. 3, Table 
R-14. Use corrected tons to determine the total 
gpm required. 



258 PRINCIPLES OF REFRIGERATION 

3. Determine the temperature differences 
between the condensing temperature and the 
"water in" and "water out" temperatures and 
find the METD by referring to Table 11-1. 

4. Make preliminary selection of condenser 
shell diameter by reference to Table R-14, 
basing the selection on the corrected tons found 
in step 1 . Find the number of tubes per pass and 
then by referring to step 2, find the gpm per 
tube. 

5. Select the desired scale factor by reference 
to Table 14-1 which suggests scale factors for 
various types of water. The most commonly 
used factor is 0.0005 and it should be borne in 
mind when selecting a factor that a determina- 
tion is being made of the frequency of cleaning 
which will be required. 

6. Referring to Fig. 1, Table R-14, determine 
the rate of heat transfer "£/" for the gpm per 
tube in step 4 and the scale factor in step 5. 

7. Calculate the surface required by use of the 
following formula. 

Square feet of surface 

Corrected tons x 14,400 
= U x METD 

8. Select a condenser having at least the 
required surface from Table R-14. Be sure to 
use the shell diameter determined in the pre- 
liminary selection of step 4. 

9. Make final checks on selection. 

a. Using the gpm per tube from step 4 and 
the nominal tube length shown in Table R-14 
for the model selected in step 8, refer to Fig. 
4 of Table R-14 to obtain water pressure drop 
through condenser. 

b. Obtain nominal operating charge from 
the last column of Table R-14. This is the 
maximum weight of liquid refrigerant which 
can be allowed in the shell during the operat- 
ing period covering some of the lower tubes. 
Larger shell diameters or separate receivers 
may be used where greater storage capacity is 
needed during operation. 

c. Determine the pump down capacity 
from Table R-14. If less than the total weight 
of refrigerant to be used in the system and 
provision for complete pump-down are re- 
quired, an additional receiver should be used. 

Example 14-9. Select an R-12 condenser 
to meet the following conditions: 



Refrigeration load 30 tons 

Condensing temperature 100° F 

Suction temperature 30° F 
Water available 2 gpm/ton 

river water reasonably clean at 78°F 

Maximum tube length 12 ft 

Maximum water pressure drop 7.5 psi 

Solution 

1. From Fig. 2, the correction factor for 
30° F suction temperature and 100° F con- 
densing temperature is 1.013. 

Corrected tons 30 x 1.013 = 30.4 tons 

2. From Fig. 3, for 2 gpm/ton the water 
temperature rise is found to be 14.4°. 

Total gpm 30.4 x 2 = 60.8 
Water "out" temperature 78 plus 14.4 

= 92.4° F 

3. GTD 100 - 78 = 22° 
LTD 100 - 92.4 = 7.6° 

From Table 11-1, METD = 13.55° F 

4. Refer to Table R-14. Use of four passes 
will usually give an economical selection for 
75° F water in and 95° F water out which 
approximates the required water conditions. 
Note that a lOf shell will probably be needed. 
This shell has sixty tubes. 

Total gpm x number of passes 

gpm per tube — — ^ — - — ; — : f 

or r Number of tubes in condenser 

_ 60.8 x 4 

60 

= 4.05 gpm per tube 

5. Referring to Table 14-1, for clean river 
water and over 3 fpm velocity, the suggested 
scale factor is 0.001. 

6. From Fig. 1, the V factor for 4.05 gpm per 
tube and 0.001 scale factor is 121.5 Btu per hour 
per square foot of extended surface per °F 
METD. 

7. Square feet required 

Corrected tons x 14,400 



U factor x METD 

_ 30.4 x 14,400 
~ 121.5 x 13.55 



= 266 sq ft 



8. Referring to Table R-14, a Model STF- 
1010 has 289 sq ft external tube surface and 
should be selected. When installed the water 
connection should be made for four-pass 
operation. 

9. (a). For water pressure drop, refer to Fig. 
4 and note that the pressure drop for 4.05 gpm 
per tube in an STF-1010 condenser connected 
for four passes is 7.1 psi. (b). Table R-14 shows 
a nominal operating charge of 38 lb of R-12, 
which will normally be sufficient for a 30-ton 



installation. However, if more operating 
storage is needed, a separate receiver may be 
chosen, or alternately a different condenser 
selection may be made if more economical, (c). 
Table R-14 also shows pump-down capacity 
which is 252 lb of R-12. Usually this will be 
sufficient, but if greater pump-down capacity is 
required, a separate receiver tank must be used. 

14-13. Simplified Ratings. Simplified ratings, 
based on the horsepower of the compressor 
driver, are available for most air-cooled and 
water-cooled condensers, particularly in smaller 
sizes. Since the power required by the com- 
pressor varies with both the evaporator load and 
the compression ratio, it provides a reasonable 
index of the condenser load at all operating 
conditions. Table R-l 5, which applies to double- 
tube condensers of the type shown in Fig. 14-12, 
is a typical simplified condenser rating table. 
14-14. Cooling Towers. Cooling towers are 
essentially water conservation or recovery de- 
vices. Warm water from the condenser is 
pumped over the top of the cooling tower from 
where it falls or is sprayed down to the tower 
basin. The temperature of the water is reduced 
as it gives up heat to the air circulating through 
the tower. 

Although there is some sensible heat transfer 
from the water to the air, the cooling effect in a 
cooling tower results almost entirely from the 
evaporation of a portion of the water as the 
water falls through the tower. The heat to 
vaporize the portion of water that evaporates is 
drawn from the remaining mass of the water so 
that the temperature of the mass is reduced. 
The vapor resulting from the evaporating pro- 
cess is carried away by the air circulating 
through the tower. Since both the temperature 
and the moisture content of the air are increased 
as the air passes through the tower, it is evident 
that the effectiveness of the cooling tower 
depends to a large degree on the wet bulb tem- 
perature of the entering air. The lower the wet 
bulb temperature of the entering air, the more 
effective is the cooling tower. 

The efficiency of a cooling tower is influenced 
by all the factors governing the rate at which the 
water will evaporate into the air (see Section 
4-8). Some of the factors which determine 
cooling tower efficiency are: (1) the mean 
difference in vapor pressure between the air and 
the water in the tower, (2) the amount of 



CONDENSERS AND COOLING TOWERS 259 

exposed water surface and the length (time) of 
exposure, (3) the velocity of the air passing 
through the tower, and (4) the direction of the 
air flow with relation to the exposed water 
surface (parallel, transverse, or counter). 

For any given water temperature entering the 
tower, the vapor pressure difference is essentially 
a function of the wet bulb temperature of the 
entering air. In general, the lower the entering 
wet bulb temperature, the greater the vapor 
pressure differential and the greater the tower 
capacity. 

The exposed water surface includes: (1) the 
surface of the water in the tower basin, (2) all 
wetted surfaces in the tower, and (3) the com- 
bined surface of the water droplets falling 
through the tower. 

Theoretically, the lowest temperature to 
which the water can be cooled in a cooling 
tower is the wet bulb temperature of the entering 
air, in which case the water vapor in the leaving 
air will be saturated. In actual practice, it is not 
possible to cool the water to the wet bulb tem- 
perature of the air. In most cases, the tem- 
perature of the water leaving the tower will be 
7° to 10° F above the wet bulb temperature of 
the entering air. Too, the air leaving the tower 
will always be somewhat less than saturated. 

The temperature difference between the tem- 
perature of the water leaving the tower and the 
wet bulb temperature of the entering air is 
called the tower "approach." As a general rule, 
all other conditions being equal, the greater the 
quantity of water circulated over the tower the 
closer the leaving water temperature approaches 
the wet bulb temperature of the air. However, 
the quantity of water which can be economically 
circulated over the tower is somewhat limited by 
the power requirements of the pump. 

The temperature reduction experienced by the 
water in passing through the tower (the differ- 
ence between the entering and leaving water 
temperatures) is called the "range" of the tower. 
Naturally, to maintain equilibrium in the con- 
denser water system, the tower "range" must 
always be equal to the temperature rise of the 
water in the condenser.* 

The load on a cooling tower can be approxi- 
mated by measuring the water flow rate over the 

* Except where a condenser by-pass is used. See 
Section 14-17. 



260 PRINCIPLES OF REFRIGERATION 

tower and the entering and leaving water tem- 
peratures. The following equation is applied : 

Tower load(Btu/min) = flow rate(gpm) 
x 8.33 x (entering water temperature 
— leaving water temperature) (14-12) 

Example 14-10. Determine the approxi- 
mate load on a cooling tower if the entering and 
leaving water temperatures are 96° F and 88° F, 
respectively, and the flow rate of the water over 
the tower is 30 gpm. 



Solution. Applying 
14-12, the tower load 
(Btu/min) 



= 30 x 8.33 

x (96 - 88) 
= 2000 Btu/min 



Since the load on the tower is equal to the 
load on the condenser, the approximate refrig- 
erating capacity of the system can be computed 
by dividing the tower load by the condenser load 
in Btu/min/ton corresponding to the operating 
conditions of the system. 

Example 14-1 1. Compute the refrigerating 
capacity of an R-13 system operating on the 
cooling tower of Example 14-10, if the evaporat- 
ing and condensing temperatures are 20° F and 
110° F, respectively. 

Solution. From Fig. 14-1, 
the load on the condenser 

= 247 Btu/min/ton 

The approximately refrig- 
erating capacity of the 
system 

Tower load (Btu/min) 



Condenser load (Btu/min/ton) 
2000 



247 
= 8.1 tons 

Since the heat absorbed per pound of water 
evaporated is approximately 1000 Btu, assuming 
a condenser load of 250 Btu/min/ton, the 
quantity of water evaporated per ton of refrig- 
eration (evaporator) is approximately 0.25 lb 
per minute or 2 gal per hour. 

In addition to the water lost by evaporation, 
water is lost from the cooling tower by "drift" 
and by "bleed-off." A small amount of water in 
the form of small droplets is entrained and 
carried away by the air passing through the 
tower. Water lost in this manner is called the 
drift loss. The amount of drift loss from a tower 



depends on the design of the tower and the wind 
velocity. 

"Bleed-off" is the continuous or intermittent 
wasting of a certain percentage of the circulated 
water in order to avoid a build-up in the con- 
centration of dissolved mineral solids and other 
impurities in the condenser water. Without 
bleed-off the concentration of dissolved mineral 
solids in the condenser will build up quite rapidly 
as a result of the evaporation taking place in the 
cooling tower. Since the scaling rate is propor- 
tional to the quality of the water, as the concen- 
tration of mineral solids in the water increases 
the scaling rate also increases. 

The amount of bleed-off required to maintain 
the concentration of dissolved mineral solids at 
a reasonable level depends upon the cooling 
range, the water flow rate, and the initial water 
conditions. Suggested bleed-off rates for various 
cooling ranges are given in Table 14-2. To 
determine the quantity of water loss by bleed- 
off, multiply the water flow rate over the tower 
by the factor obtained from Table 14-2. 

Example 14-12. Determine the quantity of 
water lost by bleed-off if the water flow rate over 
the tower is 30 gpm and the range is 10° F. 

Solution. From 
Table 14-1, the percent 
bleed-off required = 0.33 % 

The quantity of 
water lost by bleed-off = 30 gpm x 0.0033 

= 0.099 gpm 

The bleed-off line should be located in the hot 
water return line near the top of the tower so 
that water is wasted only when the pump is 
running (Fig. 14-8). 

Make-up water, to replace that lost by evapo- 
ration, drift, and bleed-off, is piped to the 
tower basin through a float valve which tends to 
maintain a constant water level in the basin. 
14-15. Cooling Tower Design. According to 
the method of air circulation, cooling towers are 
classified as either natural draft or mechanical 
draft. When air circulation through the tower 
is by natural convection, the tower is called a 
natural draft or atmospheric tower. When air 
circulation through the tower is by action of a 
fan or blower, the tower is called a mechanical 
draft tower. Mechanical draft towers may be 
further classified as either induced draft or 
forced draft, depending on whether the fan or 



Hot water in 




Cold water out 



Make-up water 
from city main 



Fig. 14-13. Natural draft-cooling tower. 

blower draws the air through the tower or forces 
(blows) it through. A schematic diagram of a 
spray-type natural draft tower is shown in Fig. 
14-13. Schematic diagrams of induced draft and 
forced draft towers are shown in Figs. 14-14 and 
14-15, respectively. 

In the spray-type atmospheric tower, the 
warm water from the condenser is pumped to 
the top of the tower where it is sprayed down 
through the tower through a series of spray 
nozzles. Since the amount of exposed water 
surface depends primarily on the spray pattern, 
a good spray pattern is essential to high effi- 
ciency. The type of spray pattern obtained 
depends on the design of the nozzles. For most 
nozzle designs, a water pressure drop of 7 to 
101b per square inch will produce a suitable 
spray pattern. 

Some natural draft towers contain decking or 
filling (usually of redwood) to increase the 



CONDENSERS AND COOLING TOWERS 261 

amount of wetted surface in the tower and to 
break up the water into droplets and slow its 
fall to the bottom of the tower. Atmospheric 
towers containing decking are called "splash- 
deck." Often, in splash-deck towers, no spray 
nozzles are used and the water is broken up into 
droplets by the "splash-impact" method. 

The quantity and velocity of the air passing 
through a natural draft cooling tower depend 
on the wind velocity. Hence, the capacity of a 
natural draft tower varies with the wind velocity, 
as does the amount of "drift" experienced. Too, 
natural draft towers must always be located out- 
of-doors in places where the wind can blow 
freely through the tower. In commercial appli- 
cations, roof installations are common. 

Since air circulation through mechanical 
draft towers is by action of a fan or blower, 
small mechanical draft towers can be installed 
indoors as well as out-of-doors, provided that 
an adequate amount of outside air is ducted into 
and out of the indoor location. Too, since 
larger air quantities and higher velocities can be 
used, the capacity of a mechanical draft tower 
per unit of physical size is considerably greater 
than that of the natural draft tower. In addition, 
most mechanical draft towers contain some sort 
of decking or fill to improve further the effi- 
ciency. Spray eliminators must be used in 
mechanical draft towers to prevent excessive 
drift losses. 

14-16. Cooling Tower Rating and Selection. 
Table R.-16 contains rating data for the spray- 
type, natural draft cooling tower illustrated in 
Fig. 14-13 and is a typical cooling tower rating 
table. Notice that the tower ratings are given 
in tons, based on a heat transfer capacity of 250 



Water in 



Water 
distributor 



Fig. 14-14. Small Induced 
draft-cooling tower. 




"Air out 



262 PRINCIPLES OF REFRIGERATION 

Air out 



\ \ M f t f t t t f, 



I Spray 
>Teliminators 



Wood fill 




Fig. 14-15. Forced draft- 
cooling tower. 



Air in 



Water out 



Btu/min/ton. Nominal tower ratings are based 
on a 3 mi per hour wind velocity, and 80° F 
design wet bulb temperature, and a water flow 
rate over the tower of 4 gpm per ton. Tower 
performance at conditions other than those 
listed in the table can be determined by using 
the rating correction chart that accompanies the 
table. 

To select the proper tower from the rating 
table, the following data must be known: 

1 . Desired tower capacity in tons (compressor 
capacity) 

2. Design wet bulb temperature 

3. Desired leaving water temperature (con- 
denser entering water temperature or tower 
approach) 

or 

1. Desired flow rate over the tower (gpm) 

2. Design wet bulb temperature 

3. Desired entering and leaving water tem- 
peratures (tower cooling range and tower 
approach) 

Example 14-13. From Table R-16, select a 
cooling tower to meet the following conditions: 

1. Required tower capacity =20 tons 

2. Design wet bulb 

temperature = 78° F 

3. Desired leaving water 

temperature = 86° F 

Solution. From Table R-16, select tower, 
Model #CSA-66, which has a capacity of 20.7 
tons at the desired conditions when the flow 
rate over the tower is 3 gpm per ton. Hence, for 
20-tons capacity, a total of 60 gpm (20 x 3) 
must be circulated over the tower. As shown in 
the table, the entering water temperature will be 
approximately 96° F. 



Exam pie 1 4- 1 4. It is desired to cool 90 gpm 
from 96° F to 86° F when the design wet bulb is 
78° F. Select the proper tower from Table 
R-16. 



Solution 
Tower range 

Tower approach 

From rating correc- 
tion chart, range-ap- 
proach factor 

From wet bulb cor- 
rection chart, wet bulb 
factor 

Nominal gpm 

From Table R-16, for 
select tower, Model #SA-68 



= 96 - 86 = 10° 

= 86 - 78 = 8° 

= 1.1 



= 1.04 

= 90 x 1.1 x 1.04 

= 103 gpm 

103 gpm nominal, 



Example 14-15. It is required to cool water 
for 30 tons at 5 gpm/ton to a 5° F approach of 
an 80° F wet bulb. Select the proper tower from 
Table R-16. 



Solution 

Total gpm required 
for 30 tons at 5 gpm/ 
ton 

From rating cor- 
don chart, rating cor- 
rection factor for 5 
gpm/ton and 5° ap- 
proach 

From wet bulb 
correction chart, wet 
bulb correction factor 

Nominal gpm 



30 x 5 = 150 gpm 



= 1.15 



1.0 

150 x 1.15 x 1.0 

172.6 gpm 



From Table R-16, for 172.6 gpm nominal, 
select tower Model #SA-612. 
14-17. Condenser By-Pass. For any given 
tower range and approach, the entering and 
leaving water temperatures will depend only on 
the wet bulb temperature of the air. Hence, in 
regions (particularly coastal areas) where the 
outdoor wet bulb temperature is relatively high, 
a closer approach to the wet bulb temperature is 
required in order to maintain a reasonable con- 
densing temperature with an economical con- 
denser size than in areas where the wet-bulb 
temperature is lower. It has already been shown 
that, in general, the greater the quantity of water 
circulated over the tower per unit of capacity 
the closer the leaving water temperature will 
approach the wet bulb temperature. Therefore, 
in regions having a high wet bulb temperature, 
it is usually desirable to circulate a greater 
quantity of water over the tower than can be 
economically circulated through the condenser 
because of the excessive pumping head encoun- 
tered. This can be accomplished by installing 
a condenser by-pass line as shown in Fig. 14-8. 
Through the use of a condenser by-pass, a 
certain, predetermined portion of the water 
circulated over the tower is permitted to by-pass 
the condenser, thereby reducing the over-all 
pumping head. 

The advantage of the condenser by-pass is 
that it makes possible the maintenance of 
reasonable condensing temperatures with mod- 
erate condenser and tower sizes without greatly 
increasing the pumping head. The quantity of 
water flowing through the by-pass is regulated 
by the hand valve in the by-pass line. Once the 
hand valve has been adjusted for the proper flow 
rate through the by-pass, the handle should be 
removed from the valve so that the valve adjust- 
ment cannot be changed indiscriminately. An 
excessive amount of water flowing through the 
by-pass will not only tend to starve the con- 
denser and raise the condensing pressure, but it 
may also cause the pump motor to become 
overloaded, thereby rendering the entire system 
inoperative. The desired flow rate through the 
by-pass is determined by subtracting the flow 
rate through the condenser from the flow rate 
over the tower. This will be illustrated presently. 

Since the cooling tower capacity must of 
necessity be equal to the condenser capacity at 
the design conditions, it follows that: 



CONDENSERS AND COOLING TOWERS 263 

Tower gpm x tower range x 500 

= condenser gpm x condenser rise x 500 
Eliminating the constant, 

Tower gpm x tower range 

= condenser gpm x condenser rise (14-13) 

Example 14-16. A compressor on a refrig- 
erating system has a capacity of 25 tons. The 
design wet bulb temperature is 80° F. The 
desired condenser water entering temperature is 
87° F and the desired temperature rise through 
the condenser is 10° F. Select a cooling tower 
from Table R-16 and determine: 

1. The total gpm circulated over the tower 

2. The temperature of the water entering the 
tower 

3. The tower cooling range 

4. The temperature of the water leaving the 
condenser 

5. The gpm circulated through the condenser 

6. The gpm circulated through the by-pass 

Solution. From Table R-16, tower, Model 
#SA-58 has a capacity of 25 tons at an 80° F 
wet bulb temperature and a 7° approach. This 
capacity is based on a water flow rate of 4 
gpm/ton and on a cooling range of 7.5° 
(94.5 - 87). 

Total gpm over the tower 
for 25 tons 

= 25 tons x 4 gpm/ton 
= 100 gpm 

From Table R-16, the 
tower entering water tem- 
perature 

= 94.5° F 
Tower range 

= 94.5 - 87 - 7.5° 
Water temperature leaving 
condenser 

= 87 + 10 = 97° F 

Rearranging and applying 
Equation 14-13, condenser 
gpm 

_ Tower gpm x tower range 

Condenser rise 
_ 100 x 7.5 

10 
= 75 gpm 

Gpm circulated through 
by-pass 

= Tower gpm — condenser gpm 
= 100-75 
= 25 gpm 



364 PRINCIPLES OF REFRIGERATION 




Refrigerant 
vapor in 



Refrigerant 

liquid Out 



\^J|/ / Eliminators 

Spray 
i ,, I >: ■ 



c 



1 



Condensing coil 



3 



Air in 



M-thL'-ut: 



•.'iL.'.'J 



!;-~.-I Water tank -_-_-_-_-: 




?■•.. m Ti 



Fig. 14-14. Schematic diagram of evaporative con- 
denser. 



14-18. Evaporative Condeneers. An evapo- 
rative condenser is essentially a water Conser- 
vation device and is, in effect, a condenser and 
a cooling tower combined into a single unit. A 
diagram of a typical evaporative condenser is 
shown in Fig. 14-16, 

As previously stated, both air and water are 
employed in the evaporative condenser. The 
water, pumped from the sump up to the spray 
header, sprays down over the refrigerant coils 
and returns to the sump. The air is drawn in 
From the outside at the bottom of the condenser 
by action of the blower and is discharged back 
to the outside at the top of the condenser. In 
some cases, both pump and biower are driven 
by the same motor. In others, separate motors 
are used. The eliminators installed in the air 
Stream above the spray header arc to prevent 
entrained water from being carried over into the 
blower. Art alternate arrangement, with the 
blower iocated on the entering air side oF the 
condenser, is shown in Fig. 14-17. 

Although the actual thermodynamic pro- 
cesses taking place in the evaporative condenser 
are somewhat complex., the fundamental process 
is that of evaporative cooling. Water is evapo- 
rated from the spray and from the wetted surface 



of the condenser into fhe air, the source of the 
vaporising heat being the condensing refrigerant 
in the condenser coil. 

The cooling produced is approximately 1000 
titu per pound of water evaporated. All the heat 
given up by the refrigerant in the condenser 
eventually leaves the condenser as either sensible 
heat or latent heat (moisture) in the discharge 
air. Since both the temperature and the mois- 
ture content of the air are increased as the air 
passes through the condenser, the effectiveness 
of the condenser depends, in part, on the wet 
bulb temperature of the entering air. The lower 
the wet bulb temperature of the entering air the 
more effective is the evaporative condenser. 

To facilitate cleaning and scale removal, the 
condensing coil is usually made up of bare 
rather than fumed tubing. The amount of coil 
surface used per ton of capacity varies with the 
manufacturer and depends to a large extent on 
the amount of air and water circulated. 

Generally, the capacity of the evaporative 
condenser increases as the quantity of air cir- 
culated through the condenser increases. As a 
practical matter, the maximum quantity of air 




Fig. 14-17. Cutaway view of "Dri-Fan" evaporative 

condenser. Funnel-shaped overflow drain provide] 
au (ornate bleed-off. (Courtesy/ Refrigeration Engln- 

Bering, Inc. A proprietary design of Refrigeration 
Engineering, Inc.) 



CONDENSERS AND COOLING TOWERS 265 



which can be circulated through the condenser 
is limited by the horsepower requirements of the 
fan and by the maximum air velocity that can be 
permitted through the eliminators without the 
carry over of water particles. 

The quantity of water circulated over the 
condenser should be sufficient to keep the tube 
surface thoroughly wetted in order to obtain 
maximum efficiency from the tube surface and to 
minimize the rate of scale formation. However, 
a water flow rate in excess of the amount 
required for adequate wetting of the tubes will 
only increase the power requirements of the 
pump without materially increasing the con- 
denser capacity. 

Assuming a condenser load of 15,000 Btu per 
hour per ton, the water lost by evaporation is 
approximately 15 lb (2 gal) per hour per ton 
(15,000/1000). In addition to the water lost by 
evaporation, a certain amount of water is lost by 
drift and by bleed-off. The amount of water 
lost by drift and by bleed-off is approximately 
l.S to 2.5 gal per hour per ton, depending upon 
the design of the condenser and the quality of 
water used. Hence, total water consumption 
for an evaporative condenser is between 3 and 4 
gal per hour per ton. 

Some evaporative condensers are available 
equipped with desuperheating coils, which are 
usually installed in the leaving air stream. The 
hot gas leaving the compressor passes first 
through the desuperheating coils where its tem- 
perature is reduced before it enters the con- 
densing coils. The desuperheating coils tend to 
increase the over-all capacity of the condenser 
and reduce the scaling rate by lowering the 
temperature of the wetted tubes. Too, often the 
receiver tank is located in the sump of the 
evaporative condenser in order to increase the 
amount of liquid subcooling. 
14-19. Rating and Selection of Evaporative 
Condensers. Table.R-17 is a typical evapo- 
rative condenser rating table. Notice that the 
ratings are based on the temperature difference 
between the condensing temperature and the 
design wet bulb temperature. The following 
sample selection is reprinted directly from the 
manufacturer's catalog data:* 

Example 14-17. Select an evaporative con- 
denser for the following conditions: 

* McQuay Products. 



6-ton evaporator load (Refrigerant- 12) 
20° evaporator temperature 
78° entering wet bulb temperature 
105° F condensing temperature 

Solution. Since the rating table is in terms of 
evaporator load at 40° F, it is necessary to 
correct for other evaporator temperatures by 
using a correction factor from R-17B as follows: 
Tons x evaporator correction factor 

•= Rating table tons 
Therefore, 6 x 1.05 = 6.3 tons. 

Referring to Table R-17A, the E-135F has a 
capacity of only 5.6 tons at 78° F entering wet 
bulb and 105° F condensing temperature. It 
does, however, have the required capacity of 6.3 
tons at between 105° F and 110° F condensing 
temperature. 

The compressor ratings should then be 
checked to see if the compressor originally 
selected has the required capacity at between 
105° F and 110° F condensing temperature. If 
not, it will be necessary to select the next larger 
size evaporative condenser or compressor to do 
the job. 

The next larger size evaporative condenser, 
the E-270F, has a capacity of 11.2 tons at the 
given conditions; however, the required capa- 
city of 6.3 tons will be obtained at a condensing 
temperature between 90 and 95° F. The com- 
pressor selection should then be made for these 
conditions. 

14-20. Water Regulating Valves. The water 
flow rate through a water-cooled condenser on a 
waste water system is automatically controlled 
by a water regulating valve (Fig. 14-18). The 
valve is installed on the water line at the inlet of 
the condenser and is actuated by the compressor 
discharge (Fig. 14-7). When the compressor 
is in operation, the valve acts to modulate the 
flow of water through the condenser in response 
to changes in the condensing pressure. An 
increase in the condensing pressure tends to 
collapse the bellows further and open the valve 
wider against the tension of the range spring, 
thereby increasing the water flow rate through 
the condenser. Likewise, as the condensing 
pressure decreases, the valve moves toward the 
closed position so that the flow rate through the 
condenser is reduced accordingly. Although 
the regulating valve tends to maintain the con- 
densing pressure constant within reasonable 
limits, the condensing pressure will usually be 
considerably higher during periods of peak 
loading than during those of light loading. 



Hi PRINCIPLES OF REFRIGERATION 




Flf. 14-lfl. Typical threaded-type water regulating valve. Larger sixes are available with, flange connection*. 
{a) Cross-sectional view showing principal parts, (b) Exterior view, (Courtesy Penn Contrail, Inc.)- 



When the compressor cycles off, the water 
valve remains open and water continues la flow 
through ihc condenser until the pressure in the 
condenser is reduced to a certain predetermined 
minimum, at which time the valve closes off 
completely and shuts o FT the water flow. When 
the compressor cycles on again, ihc water valve 
remains closed until the pressure in the con- 
denser builds up to the valve opening pressure, 
at which time the valve opens and permits water 
to flow Lh rough the condenser. The opening 
pressure of the valve is approximately 7 psi 
above the shut-oil pressure. 

The water valve is set for the desired shut-off 
pressure by adjusting the tension of the range 
spring. The minimum operating pressure for 
the valve, that is, the shut-off pressure, must be 
set high enough so that the valve will not remain 
open and permit water to flow through the con- 
denser when the compressor is on the off cycle. 
Since the saturation temperature of the refrig- 



erant in the condenser can never be lower than 
the ambient temperature at the condenser, the 
shut-off point of the water valve should be set at 
a saturation pressure corresponding to the maxi- 
mum ambient temperature in the summertime 
at the condenser location. Too, the shut-off 
pressure of the valve must be high enough so 
that the minimum condensing temperature in 
the wintertime is sufficiently high to provide a 
pressure differential across the refrigerant con- 
trol large enough to assure its proper operation. 

The capacity of water regu Sating valves varies 
with the size of the valve and the pressure drop 
across the valve orifice. The available pressure 
drop across the valve orifice is determined by 
subtracting the pressure drop through the con- 
denser and water piping from the total pressure 
drop available at the water main. 

Water regulating valves are usually selected 
from flow charts (Table fl-lS), In order to 
select the proper valve from the flow chart, the 



CONDENSERS AND COOLING TOWERS 267 



following data must be known: (1) the desired 
water quantity in gpm; (2) the maximum 
ambient temperature in the summertime; (3) 
the desired condensing temperature; and (4) 
the available water pressure drop across the 
valve. 

The following selection procedure and sample 
selection are reprinted directly from the litera- 
ture of the manufacturer:* 

1. Draw horizontal line across upper half of 
Flow Chart (Table R-18) through the required 
flow rate. 

2. Determine refrigerant condensing pressure 
rise above valve opening point. 

a. Valve closing point (to assure closure 
under all conditions) must be the refrigerant 
condensing pressure equivalent to the highest 
ambient air temperature expected at time of 
maximum load. Read this in psig from "Satu- 
rated Vapor Table" for refrigerant selected. 

b. Read from the same table the operating 
condensing pressure corresponding to selected 
condensing temperature. 

c. Valve opening point will be about 7 psi 
above closing point. 

d. Subtract opening pressure from operat- 
ing pressure. This gives the condensing 
pressure rise. 

3. Draw horizontal line across lower half of 
Flow Chart through this value. 

4. Determine the water pressure drop through 
the valve — this is the pressure actually available 
to force the water through the valve. 

a. Determine the minimum water pressure 
available from city mains or other source. 

b. From condensing unit manufacturer's 
tables read pressure drop through condenser 
corresponding to required flow. 

c. Add to this estimated or calculated drop 
through piping, etc., between water valve and 
condenser, and from condenser to drain (or 
sump of cooling tower). 

d. Subtract total condenser and piping drop 
from available water pressure. This is the 
available pressure drop through the valve. 

Example. 14-18. The required flow for an 
R-12 system is found to be 27 gpm. Condensing 
pressure is 125 psig and the maximum ambient 
temperature estimated at 86° F. City water 

» By permission of Penn Controls, Inc., Goshen, 
Indiana. 



pressure is 40 psig and manufacturer's table 
gives drop through condenser and accompany- 
ing piping and valves as 15 psi. Drop through 
installed piping approximately 4 psi. Select 
proper size of water regulating valve from Table 
R-18. 

Solution 

1. Draw a line through 27 gpm — see dotted 
line, upper half of Flow Chart (Table R-18). 

2. Closing point of valve is pressure of R-12 
corresponding to 86° F ambient = 93 psig. 

3. Opening point of valve is 93 + 7 = 100 
psig. 

4. Condensing pressure rise = 125 — 100 = 
25 psi. 

5. Draw line through 25 psi—see dotted line, 
lower half of Flow Chart. 

6. Available water pressure drop through 
valve = 40 - 19 = 21 psi. 

7. Interpolate just over the 20 psi curve — 
circle on lower half of Flow Chart. 

8. Draw vertical line upward from this point 
to flow line — circle on Flow Chart marks this 
intersection. 

9. This intersection falls between curves for 
1 in. and \\ in. valves. The 1J in. valve is re- 
quired. 

14-21. Condenser Controls. For reasons of 
economy, the condensing medium is circulated 
through the condenser only when the com- 
pressor is operating. Hence, common practice 
is to cycle the condenser fan and/or pump on 
and off with the compressor. This is usually 
accomplished by electrically interlocking the 
fan and/or pump circuit with the compressor 
driver circuit. Method of interlocking electrical 
circuits are discussed in Chapter 21. 

Whereas high pressure controls are always 
desirable as safety devices on any type of system, 
they are absolutely essential on all equipment 
employing water as the condensing medium in 
order to protect the equipment against damage 
from high condensing pressures and tempera- 
tures in the event that the water supply becomes 
restricted or is shut-off completely. The high 
pressure control has already been discussed in 
Section 13-13. 

If a refrigerating system is to function pro- 
perly and efficiently, the condensing temperature 
must be maintained within certain limits. As 
previously described, high condensing tempera- 
tures cause losses in compressor capacity and 
efficiency, excessive power consumption, and, 



268 PRINCIPLES OF REFRIGERATION 



in some cases, overloading of the compressor 
driver and/or serious damage to the compressor 
itself. 

An abnormally low condensing temperature, 
on the other hand, will cause an insufficient 
pressure differential across the refrigerant con- 
trol (condensing pressure to vaporizing pres- 
sure), which reduces the capacity of the control 
and results in starving of the evaporator and 
general unbalancing of the system. 

As a general rule, low condensing tempera- 
tures result from either one or both of two 
principal causes: (1) low ambient temperatures 
and (2) light refrigerating loads. Naturally, the 
problem of low condensing temperatures is more 
acute in the wintertime when the ambient 
temperature and the refrigerating load are both 
apt to be low. 

To maintain the condensing temperature at a 
sufficiently high level, it is necessary to make 
some provision for reducing or controlling the 




Air-cooled 
condenser 



From 

condenser 

outlet 




capacity of the condenser during periods when 
the ambient temperature is low and/or the refrig- 
erating load is light. Although the methods 
employed to control the capacity of the con- 
denser vary somewhat with the type of con- 
denser used, all involve reducing either the 
quantity of condensing medium circulated or 
the amount of effective condensing surface. 
Condenser capacity control devices are usually 
actuated by pressure or temperature controls 
which respond to condensing pressure or 
temperature. 

With regard to air-cooled condensers, the 
condensing temperature is maintained within 
the desired limits by varying the air quantity 
through the condenser or by causing a portion 
of the condenser to become filled with liquid so 
as to reduce the amount of effective condensing 
surface. 

The air quantity through the condenser is 
varied by cycling the fan or blower or by the use 

Modulating 

control valve 

(open on drop 

in pressure) 



Fig. 14-19. Winterstat control 
of air-cooled condensers, (a) 
Loop Winterstat may be used 
wherever 3 feet of head room 
is available above the top of the 
condenser. This type is the 
simplest and lowest in cost. 
(b) No-loop Winterstat is 
employed where head room 
is not available above con- 
denser. Valves and W are 
supplied as an integral unit and 
must be mounted at the level 
of the liquid outlet of the 
condenser. (The Winterstat is 
a proprietary design of the 
Kramer Trenton Company and 
is manufactured under the 
following patent numbers: 
2,564,310; 2,761,287; and 
2,869,330.) 



Constant 
inlet pressure 
throttling valve 



To receiver 



(b) 



CONDENSERS AND COOLING TOWERS 269 



Fig. 14-20. Pressure stabilizer. 
(A proprietary design of 
Dunham-Bush, Inc.) (Courtesy 
Dunham-Bush, Inc.) 



of dampers placed in the air stream. Because it 
tends to cause large fluctuations in the con- 
densing temperature, cycling of the fan cannot 
be recommended as a means of controlling the 
capacity of air-cooled condensers. Modulating 
dampers installed in the air stream provide 
satisfactory control of the air quantity in many 
cases, although some difficulty is experienced 
with dampers when the condenser is exposed to 
high wind velocities. 

A more satisfactory method of controlling 
the capacity of air-cooled condensers is to vary 
the amount of effective condensing surface by 
causing the liquid refrigerant to back up into 
the lower portion of the condenser whenever the 
condensing pressure drops below the desired 
minimum. To accomplish this, one design of 
capacity control employs a modulating valve 
installed in a by-pass line between the inlet and 
outlet of the condenser (Fig. 14-19). As the 
receiver pressure falls, the modulating valve 
opens and allows high-pressure vapor from the 
compressor discharge to flow through the by- 
pass line, thereby restricting the flow of liquid 
refrigerant from the condenser and causing the 
liquid to back up into the lower portion of the 
unit. The amount of discharge vapor by-passed, 
and therefore the amount of liquid refrigerant 
retained in the lower portion of the condenser, 
is automatically controlled by the modulating 
valve and depends upon the receiver tank 
pressure. 

Another device used to restrict the amount of 
effective condensing surface is called a "pressure 
stabilizer" (Fig. 14-20). The following descrip- 
tion of the operation of the pressure stabilizer 
is reprinted directly from the manufacturer's 
engineering data.* 

The pressure stabilizer is a heat transfer 
surface which transfers the heat from the hot gas 
discharge of the compressor to the subcooled 
liquid leaving the condenser. This heat exchange 

* Courtesy Dunham-Bush, Inc. 




is controlled by the regulating valve installed 
between the condenser and the receiver. This 
valve is set at the desired operating pressure, and 
throttles from the open position to the closed 
position as the condensing pressure drops. The 
throttling action backs up the liquid in the con- 
denser, thus reducing the amount of effective 
condensing surface. The subcooled liquid 
coming from the condenser is forced through 
the heat exchanger portion of the pressure 
stabilizer and receives enough heat from the hot 
gas to satisfactorily establish the balanced 
pressure temperature relationship in the receiver. 
This assures satisfactory condensing pressure 
and a solid column of liquid at the refrigerant 
control. 

The pressure stabilizer is designed with a pre- 
determined pressure drop to insure against 
liquid refrigerant reheating during warm weather 
operations. During high ambient air tempera- 
tures, where the condensing temperature is above 
the setting of the regulating valve, the liquid 
flows through the valve, which is fully open, and 
thereby by-passes the heat exchanger section 
(Fig. 14-21a). In Fig. 14-216, as the ambient 
temperature drops to 50° F the condensing tem- 
perature drops below the setting of the regulat- 
ing valve. The valve then modulates toward 
the closed position, and this action limits the 
flow of liquid through the regulating valve. 
Consequently, the liquid backs up in the con- 
denser until the condensing surface is reduced 
approximately 60%. The liquid which is forced 
to pass through the heat exchanger section is 
then heated up to the saturation temperature. 

When the ambient temperature drops to 0° F 
(Fig. 14-2 lc), the regulating valve throttles to 
hold 120 psi in the condenser. The liquid logs 
in the condenser so that approximately 10% of 
the surface is utilized to condense the hot gas. 

With regard to evaporative condensers, capa- 
city control is best obtained through regulation 
of the air quantity through the condenser, which 
can be accomplished either by cycling the blower 



270 PRINCIPLES OF REFRIGERATION 



90° Amb., R-12 
110* Cond. temp. 



I r' '■ '■"■ i -' vr ■■■ 

*' -■;'■■ 'v ) 

136 g ^ ' ■' 

PTTX. ........ ....'.j 




50* Amb., R-12 
102* Cond. temp. 



[ ■;■ ■ •■— 

l..."..'.'..-"l^J,»l 

,II'I.H'..I| ^T"f J 




0* Amb., R-12 
102* Cond. temp. 

- ffi'A'.",;.',..',-.'.'M 




Fig. 14-21. Air-cooled condenser control employing 
pressure stabilizer. (Courtesy Dunham-Bush, Inc.) 



or by installing dampers in the air stream. Of 
the two methods, the latter is usually the most 
satisfactory, especially where modulating dam- 
pers are used and the air quantity can be varied 
through a wide range. 



Cycling of the pump as a means of controlling 
the capacity of an evaporative condenser cannot 
be recommended. Each time the pump cycles 
off a thin film of scale is formed on the con- 
denser tubes. Consequently, frequent cycling of 
the condenser pump greatly increases the scaling 
rate, which reduces the efficiency of the con- 
denser and greatly increases maintenance costs. 
With reference to water-cooled condensers, 
recall that for a given load and condensing 
surface, the condensing temperature varies with 
the quantity and temperature of the water 
entering the condenser. Where waste water is 
used, the modulating action of the water- 
regulating valve controls the water flow rate 
through the condenser and maintains the con- 
densing temperature above the desired minimum 
so that low condensing temperatures are not 
usually a problem with waste water systems. On 
the other hand, since the flow rate of the water 
through the condenser on a recirculating water 
system is maintained constant, the condensing 
temperature decreases as the temperature of the 
water leaving the tower decreases. Therefore, 
when the ambient air temperature is low, the 
condensing temperature will also be low unless 
some means is provided for restricting the flow 
rate through the condenser or for increasing the 
temperature of the water leaving the tower. 

One method of controlling the condensing 
temperature in a recirculating water system is to 
install a water-regulating valve in the water line 
at the inlet to the condenser. The modulating 
action of the water valve will restrict the water 
flow rate through the condenser in response to a 
drop in the condensing pressure. When a water- 
regulating valve is used in a recirculating water 
system, the pressure drop through the valve 
must be taken into account in computing the 
total pumping head.* 

Where mechanical draft cooling towers are 
used, the condensing temperature can be main- 
tained at the desired level through regulation of 
the tower leaving water temperature. As in the 
case of the evaporative condenser, this can be 
accomplished by cycling the tower fan or by 
installing dampers in the air stream. 

* Except in those cases where they have a specific 
function, water-regulating valves should never be 
used in recirculating water systems, since they tend 
to restrict the water flow and increase the pumping 
head unnecessarily. 



CONDENSERS AND COOLING TOWERS 271 



14-22. Winter Operation. When the com- 
pressor and/or condenser are so located that they 
are exposed to low ambient temperatures, the 
pressure in these parts may fall considerably 
below that in the evaporator during the com- 
pressor off-cycle. In such cases, the liquid 
refrigerant, which otherwise would remain in 
the evaporator, very often tends to migrate to 
the area of lower pressure in the compressor and 
condenser. With no liquid refrigerant in the 
evaporator, an increase in evaporator tempera- 
ture is not reflected by a corresponding increase 
in the evaporator pressure, and, where the sys- 
tem is controlled by a low pressure motor 
control, the rise in evaporator pressure may not 
be sufficient to actuate the control and cycle the 
system on in response to an increase in the 
evaporator temperature. 

Corrective measures are several. One is to 
install a thermostatic motor control in series 
with the low pressure control. The thermostat 
is adjusted to cycle the system on and off, 
whereas the low pressure control serves only as a 
safety device. Another, and usually more prac- 
tical, solution is to isolate the condenser during 
the off-cycle. One method of isolating the con- 
denser during the off-cycle is illustrated in Fig. 
14-22. The check valve (Q in the condenser 
liquid line prevents the refrigerant from boiling 
off in the receiver and backflowing to the con- 



Air-cooled 
condenser 



Modulating valve 

(open on rise of 

inlet pressure) 



ggB%g 



Check valve 



To liquid_ 
receiver 



Modulating by-pass 

valve-open of drop 

in outlet pressure 




From _y 
discharge 



Fig. 14-22. Sure-start WintersUt provides normal 
head and receiver pressures when the compressor 
starts by allowing the compressor to impose its full 
discharge pressure on the liquid through the open 
(W) valve. When the receiver pressure is up to 
normal, the (R) valve opens and allows the discharge 
gas to flow to the condenser. (Courtesy Kramer 
Trenton Company.) 



4 



^Modulating 
dampers 



fa 



Water bleeds 



E 



Drain- 



Float valve 



Overflow 



rm\ 



1 



a 



Sump tank 



•k 



Pump 

Fig. 14-23. Evaporative condenser equipped with 
modulating dampers for capacity control. Protected 
auxiliary pump is designed to prevent freezing during 
winter operation. (Courtesy Refrigeration Engineer- 
ing Inc.) 

denser during the off-cycle. The (R) valve, 
which closes on drop of pressure at the valve 
inlet, closes when the compressor stops, pre- 
venting the flow of refrigerant from the evapo- 
rator, through the compressor valves and 
discharge line, into the condenser. With the 
condenser isolated, the evaporator pressure can 
build up and start the compressor regardless of 
the ambient temperature at the condenser. 

Another and rather obvious problem con- 
cerning the operation of evaporative condensers 
and cooling towers in the wintertime is the 
danger of freezing when the equipment is 
exposed to freezing temperatures. In general, 
the measures employed to prevent freezing are 
similar to those used to prevent low condensing 
temperatures, that is, controlling the air quantity 
through the tower by the use of dampers or by 
cycling the fan. In addition, an auxiliary sump 
must be installed in a warm location and the 
piping arranged so that the water drains by 
gravity into the auxiliary sump and does not 
remain in the tower or condenser sump (Figs. 
14-23 and 14-24). 

14-23. Condenser and Tower Maintenance. 
As a general rule, air-cooled condensers require 
little maintenance other than regular lubrication 
of the fan and motor bearings. However, the 



272 PRINCIPLES OF REFRIGERATION 



j ^ ** ^^^^^^™ 



Tower static 
head 



I 



Additional 
static head 



Indoor 
tank 




Fig. 14-24. Protected indoor tank. 

fan blades and condensing surface should be 
inspected occasionally for the accumulation of 
dust and other foreign materials. These parts 
should be kept clean in order to obtain high 
efficiency from the condenser. 

Any type of condenser employing water is 
subject to scaling of the condenser tubes, corro- 
sion, and the growth of algae and bacterial 
slime on all wetted surfaces. The latter is con- 
trolled by frequent cleaning of the infected parts 
and by the use of various algaecides which are 
available commercially. 

As previously stated, the scaling rate depends 
primarily upon the condensing temperature and 
the quality of water used. The scaling rate will 
be relatively low where the condenser leaving 
water temperature is below 100° F. Too, the 
importance of providing for the recommended 
amount of bleed-off cannot be overemphasized 
with regard to keeping the scaling rate at a 
minimum. In addition, a number of chemical 
companies have products which when added to 
the sump water considerably reduce the scaling 
rate. 

Scale can be removed from the condenser 



tubes by applying an approved inhibited acid 
compound, many of which are available in 
either liquid or powder form. After the tower 
or condenser sump has been drained, cleaned, 
and filled with fresh water, the cleaning com- 
pound can be added directly to the sump water. 
The pump is then started and the cleaner is 
circulated through the system until the system is 
clean, at which time the sump is again drained, 
flushed, and filled with clean water before the 
system is placed in normal operation. 

It should be pointed out that descaling com- 
pounds have an acid base and should not be 
allowed to contact grass, shrubs, or painted 
surfaces. Therefore, it is usually advisable to 
remove the cooling tower spray nozzles, if any, 
in order to minimize the danger of damaging 
shrubs or painted surfaces with drift from the 
tower. 

When rapid descaling of the condenser tubes 
is required, an inhibited solution (18%) of 
muriatic acid may be used. However, muriatic 
acid should be used only on the condenser tubes. 
The system pump should not be used to circu- 
late the acid. A small pump having an acid 
resistant impeller (brass or nylon) may be used 
for this purpose (see Fig. 14-25). After the 
condenser is clean, it should be flushed with 
clean water or with an acid neutralizer as 
recommended by the manufacturer. 

Corrosion is usually greatest in areas near 
salt water or in industrial areas where relative 



Condenser - 




Fig. 14-25. Apparatus for descaling condenser. 



large concentrations of sulfur and other indus- 
trial fumes are found in the atmosphere. Corro- 
sion damage is minimized by regular cleaning 
and painting of the affected parts and by 
application of protective coatings of various 
types. 

PROBLEMS 

1. An R-12 system is operating at an evaporator 
temperature of 0° F and a condensing tempera- 
ture of 100° F. From Chart 14-1, determine the 
heat load on the condenser in Btu per minute 
per ton of refrigeration. Ans. 257 Btu/min/ton 

2. An R-22 system operating with a 40° F 
evaporator and a 110° F condenser has an 
evaporator load of 10 tons. Determine the heat 
load on the condenser in Btu/hr. 

Ans. 141,000 Btu/hr 

3. The heat rejected to a water-cooled con- 
denser is 120,000 Btu/hr. How many square 
feet of effective tube surface must this condenser 
have if the U factor of the condenser is 100 
Btu/hr/sq ft/° F and the METD is 5° F at the 
desired gpm? Ans. 240 sq ft 

4. The heat load on the evaporator of an air 
conditioning system is 60,000 Btu/hr. If the 
coefficient of performance of the system is 4 : 1, 
what is the heat load on the condenser in Btu/hr ? 

Ans. 75,000 Btu/hr 

5. An R-12 waste water system operating at a 
40° F suction temperature and a 105° F con- 
densing temperature has an evaporator load of 
5 tons. If the condenser is selected for a 12° F 
water temperature rise, how many gpm must be 
circulated through the condenser? 

Ans. 11.5 gpm 

6. Seventy-two gallons of water per minute are 
circulated through a water-cooled condenser. 
If the temperature rise of the water in the con- 
denser is 14° F, what is the heat load on the 
condenser? Ans. 504,000 Btu/hr 

7. An R-12 air conditioning system operating 
with an evaporator temperature of 40° F and a 
condensing temperature of 120° F has an 
evaporator load of 60,000 Btu/hr. 4500 cfm of 
air are circulated over the condenser. If the 
temperature of the air entering the condenser is 
90° F, compute: (a) the leaving air temperature 
and (b) the METD. 

Ans. (a) 104.6° F (b) 21.89° F 

8. If the air-cooled condenser in Problem 7 has 
a free face area of 5.5 sq ft, what is the velocity 
of the air through the condenser? 

Ans. 818 fpm 



CONDENSERS AND COOLING TOWERS 273 

9. From Table R-12, select an air-cooled con- 
denser for a compressor having a capacity of 
42,000 Btu/hr if the design suction and discharge 
temperatures are 40° F and 130° F, respectively, 
and the outdoor design dry bulb temperature 
for the region is 95° F. 

10. Select a shell-and-tube water-cooled con- 
denser for an R-12 system to meet the following 
conditions : 

Refrigeration load and eva- 
porator 60 tons 
Evaporator temperature 40° F 
Condensing temperature 1 10° F 
Water quantity 2.5 gpm/ton 
Untreated cooling tower water enters con- 
denser at 85° F. 

11. Rework Problem 10 using a condensing 
temperature of 120° F. 

12. A cooling tower and a water-cooled con- 
denser (with by-pass) are operating with a 
condenser load of 240,000 Btu/hr. Forty-eight 
gpm are circulated through the condenser and 
32 gpm are by-passed. The ambient wet bulb 
temperature is 78° F and the tower approach is 
7° F. Determine: 

(a) The temperature of the water entering the 
condenser. Ans. 85° F 

(6) The temperature of the water leaving the 
condenser. Ans. 95° F 

(c) The temperature of the water entering the 
cooling tower. Ans. 91° F 

(</) The tower range. Ans. 6° F 

13. A compressor on a Refrigerant- 12 system 
has a capacity of 50 tons. The design wet bulb 
temperature is 78° F. The desired condenser 
water entering temperature is 85° F and the 
desired temperature rise through the condenser 
is 12° F. Select a cooling tower from Table R-15 
and determine: 

(a) The total gpm circulated over the tower 

(b) The temperature of the water entering the 
tower 

(c) The temperature of the water leaving the 
condenser 

(d) The tower range 

(e) The gpm circulated through the condenser 
(/) The gpm by-passed 

14. Select an evaporative condenser for the 
following conditions: 

Refrigerant- 12 system 

Evaporator load — 10 tons 

Evaporator temperature — 40° F 

Wet bulb temperature of entering air — 78° F 

Condensing temperature — 105° F 



15 

Fluid Flow, 
Centrifugal Liquid 
Pumps, Water and 
Brine Piping 



15-1. Fluid Pressure. The total pressure 
exerted by any fluid is the sum of the static and 
velocity pressures of the fluid, viz: 

Pt =P* +Pv (15-1) 

where p t = the total pressure 
p, = the static pressure 
p v = the velocity pressure 

All flowing fluids possess kinetic energy and 
therefore exert a force or pressure in the direc- 
tion of flow. The pressure exerted by a fluid 
which is the direct result of fluid motion or velo- 
city is called the velocity pressure of the fluid. 
Any pressure exerted by a fluid which is not the 
direct result of fluid motion or velocity, regard- 
less of the force causing the pressure, is called 
the static pressure of the fluid. For fluids at rest 
(static), the velocity pressure is equal to zero 
and the total pressure is equal to the static 
pressure. Whereas velocity pressure acts only 
in the direction of flow, static pressure acts 
equally in all directions. This is easily demon- 
strated through the use of an example employing 
a gravitational column. 

It was shown in Chapter 1 that the action of 
gravity on any body causes the body to exert 
a force which is commonly referred to as the 
weight of the body. For a solid material, 
because of the rigid molecular structure, the 



gravitational force or pressure is exerted in a 
downward direction only. However, because of 
the loose molecular structure of fluids, the gravi- 
tation force or pressure exerted at any point in a 
body of fluid acts equally in all directions — up, 
down, and sideways, and always at right angles 
to any containing surfaces. When no force 
other than the force of gravity is acting on the 
fluid, the pressure at any depth in a body of fluid 
is proportional to the weight of fluid above that 
depth. When an external force in addition to 
the force of gravity is applied to the liquid, the 
pressure at any depth in the fluid is proportional 
to the weight of the fluid above that depth, plus 
the pressure caused by the external force. 

For example, assume that a flat-bottomed 
container 1 sq ft in cross section and 10 ft high 
is filled to the top with water at a temperature of 
60° F (Fig. 15-1). Since water at 60° F has a 
density of 62.4 lb per cubic foot, if the pressure 
of the atmosphere on the surface of the water is 
neglected, the total force acting on the bottom 
of the tank due to the weight of the water alone 
is 624 lb (10 x 62.4). Since the base area of the 
tank is 1 sq ft, the pressure exerted on the 
bottom of the tank is 624 psf or 4.33 psi 
(624/144). Since this pressure acts equally in all 
directions, it is exerted on the sides of the tank 
at the base as well as on the bottom of the tank. 

Assume now that level A in the water column 
is exactly 1 ft below the surface of the water. 
The volume and weight of water above this level 
are 1 cu ft and 62.4 lb, respectively. Since this 
weight of water is also evenly distributed over an 
area of 1 sq ft, the fluid pressure acting in all 
directions from any point at level A is 62.4 psf 
or 0.433 psi. Similarly, the volume and weight 
of water above level B, which is located 5 ft 
below the surface of the water, are 5 cu ft and 
3121b (5 x 62.4), respectively, and the fluid 
pressure at this level is 312 fsf or 2.165 psi. 

If the force exerted on the top of the water by 
the pressure of the atmosphere is taken into 
account, the pressure of the water at any level 
in the tank will be increased by an amount equal 
to the pressure of the atmosphere. Assuming 
normal sea level pressure, the fluid pressures at 
levels A and B are 15.129 psi (0.433 + 14.696) 
and 16.861 (2.165 + 14.696), respectively, while 
the pressure at the base of the tank is 19.026 psi 
(4.33 + 14.696). However, it should be recog- 
nized that since the pressure of the atmosphere 



274 



FLUID FLOW, CENTRIFUGAL LIQUID PUMPS, WATER AND BRINE PIPING 275 



is exerted also on the outside of the tank the 
pressure tending to burst the tank is still only 
that resulting from the gravitational effect on 
the water alone. 

For any noncompressible fluid (liquid), the 
pressure exerted by the fluid at any level in a 
fluid column is directly proportional to the 
depth of the fluid at that level.* Hence, the 
pressure of a liquid at any level in a column of 
liquid can be determined by multiplying the 
depth at that level times the density of the fluid, 
viz: 

Pressure (psf) = depth (ft) x density (lb/cu ft) 

(15-2) 



Pressure (psi) 



depth (ft) x density (lb/cu ft) 
144 



(15-3) 

15-2. Head-Pressure Relationship. The 

vertical distance between any two levels in a 
column of liquid is called the "head" of the 
liquid at the lower level with respect to the upper 
level. For example, with respect to level B in 
Fig. 1 5- 1 , the head of the water at the base of the 
column is 5 ft. With respect to the top of the 
column, the head of the water at the base of 
the column is 10 ft. Similarly, with respect to the 
top, the water heads at levels A and £ are 1 ft 
and 5 ft, respectively. 

Since the depth of the liquid at any level in a 
liquid column is equal to the head of the liquid 
at that level with respect to the top of the column, 
the head can be substituted for depth in 
Equation 15-3 and the following relationship 
between head and pressure is established: 



Pressure (psi) = 



Head (ft) x density (3/cu ft) 
144 



(15-4) 

„ T ,. ,, . Pressure (psi) x 144 
Head (ft) = — — . *, /, — — - (15-5) 
Density (lb/cu ft) K ' 

It is evident from the foregoing that there is a 
definite and fixed relationship between the head 
and the pressure of any liquid, the head-pressure 
ratio for any given liquid being dependent upon 
the density of the liquid. For example, in the 
case of water, the head-pressure ratio is 2.31 ft 

* This is not true of a compressible fluid because 
the density of a compressible fluid varies with the 
depth. 



P = 




Fig. 15-1. Illustrating head-pressure relationship. 



to 1 psi. For mercury, the head-pressure ratio 
is 2.04 in. to 1 psi. This means that a pressure 
of 1 psi is equivalent, to head of 2.31 ft of water 
column or 2.04 in. of mercury column. Con- 
versely, a 1 ft column of water (1 ft water head) 
is equivalent to 0.433 psi, whereas a 1 ft column 
of mercury (1 ft mercury head) is equivalent to 
24.48 psi. 

With respect to the head-pressure relationship, 
the following general statements can be made: 

1. For any liquid of given and uniform 
density, the pressure exerted by the liquid is 
directly proportional to the head of the liquid. 

2. At any given head, the pressure exerted by 
any liquid is directly proportional to the density 
of the liquid. Liquids having different densities 
will exert different pressures at the same head. 

15-3. Static and Velocity Heads. The total 
head of any fluid is the sum of the static and 
velocity heads of the fluid, viz: 

h t =h s + h v (15-6) 

where, h t = the total head in feet 
h, = the static head in feet 
h v = the velocity head in feet 
The static head of any liquid is expressed as 
the height in feet (or inches) of a gravitational 
column of that liquid which would be required 



276 PRINCIPLES OF REFRIGERATION 



Static 

n y pr T ure f\ |-t 



Fig. 15-2. Illustrating relationship between the 
static, velocity, and total pressures of a fluid flowing 
in a circuit. 

to produce a base pressure equal to the static 
pressure of the liquid. That is, the head in feet 
of liquid column equivalent to the static pressure 
of the liquid is called the static head of the 
liquid. Likewise, the head in feet of liquid 
column equivalent to the velocity pressure of a 
liquid is called the velocity head of the liquid. 

The fundamental relationship between velo- 
city and velocity head is established by Galileo's 
law, which states in effect that all falling bodies, 
regardless of weight, accelerate at equal rates 
and that the final velocity of any falling body, 
neglecting friction, depends only upon the height 
from which the body falls. Hence, the height in 
feet from which a body must fall in order to 
attain a given velocity is the velocity head corre- 
sponding to that velocity. The velocity head 
corresponding to any given velocity can be 
determined by applying the following equation: 



h -*" 



(15-7) 



where, h v = the velocity head in feet 

V = the velocity in feet per second (fps) 
g = the acceleration due to gravity (32.2 
ft/sec/sec) 
By combining and/or rearranging Equations 
15-7 and 15-4, the following relationships are 
established: 
To convert velocity head to velocity pressure, 

JLxi 
* - "14T (15 " 8) 

To convert velocity to velocity pressure, 
V* x P 
Pv = Ig x 144 



To convert velocity head to velocity, 

V = y/2g x h v (15-10) 

To convert velocity pressure to velocity, 



fa * /». * 

V P 



144 



(15-11) 



(15-9) 



15-4. Head-Energy Relationship. Although 
the term "head" itself is entirely independent of 
weight or density, it should be recognized that 
the head of any fluid is numerically equal to the 
energy per pound of fluid. For this reason, head 
is often used to express energy per pound of 
fluid. 

The basic relationship of head to energy or 
work is shown in the following equation : 

Energy or work (ft-lb) = mass (lb) x head (ft) 

(15-12) 

Since velocity head (h v ) is equal to V*l7g 
(Equation 15-7), it follows that the total velocity 
(kinetic) energy (E„) of any given mass (Af ) of 
fluid flowing at any given velocity (V) can be 
expressed as 

E k =Mx — 

The fact that the preceding equation is iden- 
tical to Equation 1-7 indicates that the velocity 
head of a fluid is an expression of the kinetic 
energy per pound of fluid. Similarly, it can be 
shown also that the static head of a fluid is an 
expression of the potential energy per pound of 
fluid. 

In any fluid column of uniform and constant 
density, the potential energy per pound of fluid 
is the same at all levels in the column. However, 
the potential energy at various levels is differently 
divided between the energy of position and the 
energy of pressure (head) depending upon the 
elevation. For example, in Fig. 15-1, 1 lb of 
water at the uppermost level in the tank has a 
potential energy of position with relation to the 
base of 10 ft-lb (1 lb x 10 ft) in accordance with 
Equation 1-8. Since the head at this level is 
zero, the potential energy of pressure (head) is 
also zero. On the other hand, 1 lb of water at 
the base of the tank has no potential energy of 
positions, but has pressure or head energy of 
10 ft-lb (1 lb x 10 ft), according to Equation 
15-12. Likewise, 1 lb of water at a level midway 
in the water column also has potential energy in 
the amount 10 ft-lb, the energy being evenly 



FLUID FLOW, CENTRIFUGAL LIQUID PUMPS, WATER AND BRINE PIPING 277 



divided between the energy of position and the 
energy of pressure. 

15-5. Static Head-Velocity Head Relation- 
ship in Flowing Fluids. The fact, that the 
static pressure of a fluid is exerted equally in all 
directions, whereas the velocity pressure of the 
fluid is exerted only in the direction of flow, 
makes it relatively simple to measure the static 
and velocity pressures (or heads) of a fluid 
flowing in a conduit. This is illustrated in Fig. 
15-2. Notice that tube A is so connected to the 
conduit that the opening of the tube is exactly 
perpendicular to the line of flow. Since only the 
static pressure of the fluid will act in this direc- 
tion, the height of the fluid column in tube A is 
a measure of the static pressure or static head of 
the fluid in the conduit. On the other hand, 
tube B is so arranged in the conduit that the 
opening of the tube is directly in the line of flow. 
Since both the static pressure and the velocity 
pressure of the flowing fluid act on the opening 
of tube B, the height of the liquid column in tube 
B is a measure of the total pressure or total head 
of the fluid. Since the total pressure or head of a 
fluid is the sum of the static and velocity 
pressures or heads, it follows that the difference 
in the heights of the two fluid columns is a 
measure of the velocity pressure or velocity head 
of the fluid in the conduit. 

If losses because of friction are neglected, the 
total pressure or head of a flowing fluid will be 
the same at all points along the conduit. How- 
ever, the total head may be differently divided 
between static head and velocity head at the 
several points, depending upon the velocity of 
the fluid at these points. 



For any given flow rate (quantity of flow), the 
velocity of the fluid flowing in a conduit varies 
inversely with the cross-sectional area of the 
conduit. This relationship is expressed by the 
basic equation 

(15-13) 



*-\ 



where V = the velocity in feet per second 

Q = the flow rate in cubic feet per second 
A = the cross-sectional area of the con- 
duit in square feet 

Note. When Q is in cubic feet per minute, 
V will be in feet per minute. 

In accordance with Equation 15-13, the fluid 
velocity (and velocity head) in section B of the 
conduit in Fig. 15-3 is greater than that in 
sections A and C, since the cross-sectional area 
of section B is less than that of sections A and C. 
Assuming that the total head of the fluid is the 
same at all points in the conduit, it follows then 
that the static head-velocity head ratio in section 
B is different from that in sections A and C. As 
the fluid flows through the reducer between 
sections A and B, static head is converted to 
velocity head (pressure is converted to velocity). 
Conversely, as the fluid flows through the 
increaser between sections B and C, velocity 
head is converted back into static head (velocity 
is converted to pressure). 

In view of the head-energy relationship, it is 
evident that the conversion of static head to 
velocity head is in fact a conversion of potential 
energy (pressure) into kinetic energy (velocity). 
Likewise, the conversion of velocity head to 
static head represents a conversion of kinetic 
energy (velocity) to potential energy (pressure). 




Fig. 15-3. Illustrating changes in static-velocity pressure ratio resulting from changes in conduit area. 



278 PRINCIPLES OF REFRIGERATION 



15-6. Friction Head. It has already been 
established that a fluid flowing in a conduit will 
suffer losses in energy (converted into heat) as a 
result of the work of overcoming friction. These 
energy losses are frequently expressed in terms 
of pressure drop or head loss. The pressure drop 
in psi or the head loss in feet experienced by a 
fluid flowing between any two points in a con- 
duit is known as the friction head or friction loss 
between these two points. 

The amount of pressure drop or head loss 
suffered by a fluid due to friction in flowing 
through a conduit varies with a number of 
factors: (1) the viscosity and specific gravity of 
the fluid, (2) the velocity of the fluid, (3) the 
hydraulic radius (ratio of perimeter to diameter) 
of the conduit, (4) the roughness of the internal 
surface of the conduit, and (5) the length of the 
conduit. 

Obviously, the mathematical evaluation of all 
these factors is too laborious for most practical 
purposes. As a general rule, the friction loss in 
piping is determined from charts and tables. 

The pressure (friction) loss in psi per hundred 
feet of straight pipe is given in Charts 15-1 and 
15-2 for various flow rates in various sizes of 
pipe. Chart 15-1 applies to smooth copper tube, 
whereas Chart 15-2 applies to fairly rough pipe. 
Since the pressure loss for a given pipe size and 
flow rate is proportional to the length of the 
pipe, the pressure loss through any given length 
of straight pipe is determined by the following 
equation: 

Total pressure loss (ft) 

Total length of pipe (ft) 
" 100 

x pressure loss/100 ft (psi) (15-14) 

Pipe fittings, such as elbows, tees, valves, etc., 
offer a greater resistance to flow than does 
straight pipe and therefore must be taken into 
account in determining the total friction loss 
through the piping. For convenience, this is 
done by considering the fittings as having a 
resistance equal to a certain length of straight 
pipe called the "equivalent length." Table 15-1 
lists the equivalent length of straight pipe for 
various types of fittings and valves. Notice that 
the equivalent length varies with the size of the 

Pipe- 
When the equivalent length of the fittings is 

added to the actual length of straight pipe, the 



result is called the "total equivalent length." 
This value is then applied in Equation 15-14 to 
determine the total friction loss through the 
piping. 

Example 15-1. A water piping system con- 
sists of 128 ft of 2 in. straight pipe, 6 standard 
elbows, and 2 gate valves (full open). Using 
fairly rough pipe, if the flow rate thrdugh the 
system is 40 gpm, determine: 

(a) The total equivalent length of straight pipe 

(b) The total friction loss through the piping 
in psi and in feet of water column. 

Solution. From Table 15-1, the equivalent 
lengths of 2 in. standard elbows and 2 in. gate 
valves (full open) are 5 ft and 1 .2 ft, respectively. 
From Chart 15-2, for a flow rate of 40 gpm, the 
friction loss per hundred feet of 2 in. nominal 
pipe is 3 psi. From Table 1-1, a pressure of 
1 psi is equivalent to 2.31 ft of water column. 
(a) Total equivalent length 

Straight pipe = 128.0 ft 

Six 2 in. elbows @ 5 ft = 30.0 

Six 2 in. gate valves @ 
1.2 ft 2.4 





160.4 ft 


(b) Applying Equation 
15-14, the total friction 


loo XJ 


loss through the piping 


Converting to ft H 4 


= 4.8 psi 
= 4.8 x 2.31 
= 11.09 ft H 2 



Although the pressure loss determined from 
Charts 15-1 and 15-2 apply only to water, the 
charts can be used for other fluids by multiplying 
the water pressure loss obtained from these 
charts by the correction factors listed in Table 
15-2. 

15-7. Centrifugal Pumps. Liquid pumps used 
in the refrigerating industry to circulate chilled 
water or brine, and the condenser water are 
usually of the centrifugal type. 

A centrifugal pump consists mainly of a 
rotating vane-type impeller that is enclosed 
in a stationary casing. The liquid being pumped 
is drawn in through the "eye" of the impeller 
and is thrown to the outer edge or periphery of 
the impeller by centrifugal force. Considerable 
velocity and pressure are imparted to the liquid 
in the process. The liquid leaving the periphery 
of the impeller is collected in die casing and 
directed through the discharge opening (Fig. 
15-4). 



FLUID FLOW, CENTRIFUGAL LIQUID PUMPS, WATER AND BRINE PIPING 279 



Frequently, the impeller of the pump is 
mounted directly on the shaft of the pump- 
driving motor so that the pump and motor are 
an integral unit (Fig. 1 5-5). In other cases, the 
pump and motor are separate units and are 
connected together by a. flexible coupling. 

In general, the capacity of a centrifugal pump 
depends on the design and size of the pump and 
on the speed of the motor. For a pump of 
specific size, design, and speed, the volume of 
liquid handled varies with the pumping head 




Flg> 15-4. Fluid flow through centrifugal pump. 
(Courtejy Ingereoll-iUfld Company.) 



against which the pump must wort. A charac- 
teristic head-capacity curve for a typical centri- 
fugal pump is shown in Fig. 1 5-6. Notice that 
the pumping head is maximum when the valve 
on the discharge of the pump is closed, at which 
time the pump delivery is zero. As the valve is 
opened, the pumping head decreases and the 
deliver rate increases. 

Centrifugal pumps are rated in gpm of 
delivery at various pumping heads, that is, 
centrifugal pumps are rated to deliver a certain 
gpm against a certain pumping head. Although 




Fig. 15-5, Typical centrifugal pump and motor 
asig m bl y . (Gourteiy Bell & Getsttt Co m pany ,) 



pump ratings are available in table form, more 
frequently they are taken from head-capacity 
curves (see Chart R-19). In either case, before 
the proper pump can be selected from the manu- 
facturer's ratings, it is necessary to know the 
required gpm and the total pumping head 
against which the pump must operate. 
15-8. Total Pumping Head. The total pump- 
ing head is the sum of the static head and the 
friction head. 

The static head is the vertical distance be- 
tween the "free liquid level' 1 and the highest 
point to which the liquid must be lifted by the 
pump. For the condenser-water circulating sys- 
tem in Fig. 15-7, the static bead, measured in 
feet of water column. Is the vertical distance in 
feet between the free water level in the tower 
basin and the tower spray header. Because of 
the water head in the tower basin, the water in 
the discharge pipe will stand to the level of the 




Gpm 
Fig. 15-4. Centrifugal pump dalivery i 
erewei h the pumping h«ad dttrn*«. 
IngsrsoH-Rind Company.) 



apacity in- 
{CourtMy 



280 PRINCIPLES OF REFRIGERATION 




Total length of 

straight pipe 

80 ft 



Gate valve > 



Condenser 



Static 
head 



Globe 
/^valve 




Pump-^ 
Fig. 15-7. Condenser-water circulating system. 



and cooling towers, are found in the manu- 
facturers' rating tables. 

When more than one condenser (or chiller, 
etc.) is used in the system, the condensers are 
piped in parallel and only the condenser with the 
largest pressure drop is considered in computing 
the pumping head. 

The pressure loss through the cooling tower, 
as given by the tower manufacturer, is the total 
head and includes both the tower static and 
friction heads. Therefore, the static head of the 
tower should not be considered separately in 
determining the total pumping head. When the 
tower static head is the only static head in the 
system, the static head should be disregarded 
entirely. However, in the event that an auxiliary 
indoor storage tank is employed, as shown in 
Fig. 14-24, the vertical distance between the 
level of the water in the tank and the normal 
water level in the tower basin must be treated 
as a separate static head. 

Since pump manufacturers always express the 
pumping head in "feet of water column," it is 
necessary to compute the pumping head in these 
units. When the pressure loss through the 
several system components is given in psi or in 
other units of pressure, it must be converted to 
feet of water column before it can be used in 
computing the pumping head. The required 
conversion factors are found in Table 1-1. 



water in the tower basin of its own accord. 
Therefore, the distance the water is actually 
lifted by the pump is only the distance from the 
water level in the tower basin up to the spray 
header. Contrast this with the pumping system 
shown in Fig. 15-8. 

When the piping system is a closed circuit, as 
in Fig. 1 5-9, there is no static head on the pump, 
since the fluid on one side of the piping system 
will exactly balance the fluid on the other side. 

A typical piping system curve in which gpm 
is plotted against total head is shown in Fig. 
15-10. Notice that the total head increases as 
the flow rate through the system increases and 
that the increase in the total head results 
entirely from an increase in the friction head, 
the static head being constant. 
15-9. Determining the Total Pumping 
Head. The pressure loss through the various 
system components, such as condensers, chillers, 



Static 
head 



■* 1 




Lower 
tank 



* | [|Purnp| 



Fig. 15-8 



Pipe system 
curve 



Friction loss 



FLUID FLOW, CENTRIFUGAL LIQUID PUMPS, WATER AND BRINE PIPING 281 

Example 15-2. The recirculating water sys- 
tem shown in Fig. 15-7 is for a 10-ton refrig- 
erating system. The flow rate over the tower is 
40 gpm (4 gpm/ton). The flow rate through the 
condenser is 30 gpm (3 gpm/ton), with 10 gpm 
(1 gpm/ton) flowing through the condenser by- 
pass. From the manufacturers' rating tables, 
the tower head based on 4 gpm/ton is 24 ft of 
water column, whereas toe pressure drop 
through the condenser for 30 gpm is 11.2 psi or 
25.9 ft of water column (11.2 x 2.31). If the 
size of the piping is 2 in. nominal, determine the 
total pumping head and select the proper pump 
from Chart R-19. 



s 




Static head 
(or elevation) 



Solution. Total equiva- 
lent length of pipe: 
Straight pipe 
3 — 2 in. standard el- 
bows at 5 ft 
2 — 2 in. tees (side out- 
let) at 12 ft 

4—2 in. gate valves 
(open) at 1.2 ft 

From Chart 15-2, the 
pressure loss per 100 ft of 
pipe (40 gpm and 2 in. pipe) 

Applying Equation 15-14, 
the total pressure loss 
through the piping 



Open balance 
tank "" 



80.0 ft 

15.0 

24.0 

4.8 
123.8 ft 



3 psi 
123.8 



x 3 



100 
3.71 psi 



Chilled water 

air-cooling 

coil 




Water chiller 



Fig. 15-9. Closed chilled water (or brine) circulating 
system. To compute pumping head use circuit 
having greatest friction loss. There is no static head. 



Gpm 

Fig. 15-10. Friction head of piping system increases 
as flow rate through system increases. (Courtesy 
Ingersoll-Rand Company.) 



Converting to ft H a O 

Total pumping head 
Piping 
Condenser 
Tower 



= 3.71 x 2.31 
= 8.58 ft H 2 

= 8.58 ft 
= 25.90 
= 24.00 
= 57.58 ft H 2 
From Chart R-19, select pump Model 

#1531-28, which has a delivery capacity of 40 

gpm at a 57-ft head. 

Example 15-3. At the required flow rate of 
100 gpm, a certain water system has a pumping 
head of 60 ft of water column. Select the proper 
pump from Chart R-19. 

Solution. Reference to Chart R-19 shows that 
pump Model #1531-30 is the smallest pump 
which can be used. However, since this pump 
will deliver 125 gpm at a 60-ft head, to obtain 
the desired flow rate of 100 gpm, the pumping 
head must be increased to 73 ft of water. This is 
accomplished by throttling the pump with a 
globe valve installed on the discharge side of the 
pump. (The pump should never be throttled on 
the suction side.) 

15-10. Power Requirements. The power 
required to drive the pump depends upon the 
delivery rate in pounds per minute, the total 
pumping head, and the efficiency of the pump, 
viz: 

Pounds per minute x total head in feet 
" 33,000 x pump efficiency 

Since the flow rate is usually in gpm, a more 
practical equation is 

Gpm x total head x 8.33 lb/gal 



Bhp =■ 



33,000 x efficiency 



282 PRINCIPLES OF REFRIGERATION 



Combining constants, 

Gpm x total head in feet 
Bhp " ■ 3960 x efficiency (15 - 15) 

Equation IS- IS applies to water. When a 
liquid other than water is handled, the specific 
gravity of the liquid must be taken into account, 
viz: 

Gpm x total head x specific gravity 



Bhp = 



3960 x efficiency 



(15-16) 



From Equation 15-16, it is evident that the 
power required by the pump increases as the 
delivery rate, total head, or specific gravity 
increases, and decreases as the pump increases. 

Typical pump horsepower and efficiency 
curves are shown in Fig. 15-11. Notice that 
pump horsepower is lowest at no delivery and 
increases progressively as the delivery rate 
increases. Hence, any decrease in the pumping 
head will cause an increase in both the delivery 
rate and the power requirements of the pump. 

Pump efficiency, also lowest at no delivery, 
increases to a maximum as the flow rate is 
increased and then decreases as the flow rate is 
further increased. The pump efficiency curve in 
Fig. 15-11 indicates that the highest efficiency is 
obtained when the pump is selected to deliver 
the desired gpm when operating at some point 
near the midpoint of its head-capacity curve. 
15-11. Water Piping Design. In general, the 
water piping should be designed for the mini- 
mum friction loss consistent with reasonable 
initial costs so that the pumping requirements 
are maintained at a practical minimum. Water 




Gpm 



Fig. 15-1 1. Variations in pump horsepower and 
efficiency with delivery rate. (Courtesy Ingersoll- 
Rand Company.) 



lines should be kept as short as possible and a 
minimum amount of fittings should be used. 

Standard weight steel pipe or Type "L" 
copper tubing are usually employed for con- 
denser water piping. Pipe sizes which will 
provide water velocities in the neighborhood of 
5 to 8 fps at the required flow rate will usually 
prove to be the most economical. For example, 
assume that 150 gpm of water are to be circu- 
lated through 100 equivalent feet of piping. The 
following approximate values of velocity and 
friction loss are shown in Chart 15-2 for a flow 
rate of 150 gpm through various sizes of pipe: 

Pipe Size Velocity Friction Loss per 100 ft 
(inches) (fps) (psi) (ftH a O) 

2 15.5 31.5 72.8 
2\ 10.0 10.5 24.3 

3 7.1 4.8 11.1 
3| 5.2 2.0 4.6 

4 3.9 1.1 2.5 

Notice that whereas increasing the pipe size 
from 2 to 3 in. results in a considerable reduction 
in the friction loss (from 72.8 to 11.1ft), a 
further increase in the pipe size from 3 to 4 in. 
reduces the friction loss by only an additional 
9.4 ft of water column (11.1 to 2.54 ft), which 
will not ordinarily justify the increase in the cost 
of the pipe. Depending upon the characteristics 
of the available pump, either 3 in. or 3 J in. pipe 
should be used. For instance, assume two 
separate systems having pumping heads, ex- 
clusive of tiie friction loss in the piping, of 55 ft 
of water column and 65 ft of water column, 
respectively. Reference to Chart R-19 indicates 
that the only suitable pump for either of the 
systems is Model 31531-32, which has a delivery 
rate of 150 gpm at a 70-ft head. Therefore, for 
the system having the 55-ft head, the permissible 
friction loss in the piping is 15 ft (70 — 55), 
whereas for the system having the 65-ft head, 
the permissible friction loss in the piping is only 
5 ft (70 — 65). For the latter system, 3J in. pipe 
must be used, since the use of 3 in. pipe would 
result in a total pumping head in excess of the 
allowable 70 ft and necessitate the use of the 
next larger size pump. On the other hand, for 
the former system, 3 in. pipe is the most prac- 
tical size. The use of 3 J in. pipe in this instance 
would result in a total pumping head of only 
61 ft and would necessitate throttling of the 



FLUID FLOW. CENTRIFUGAL LIQUID PUMPS, WATER AND BRINE PIPING 283 



pump discharge in order to raise the pumping 
head to 70 ft and obtain the desired flow rate of 
ISO gpm. 

In designing the piping system, care should be 
taken to include all valves and fittings necessary 
for the proper operation and maintenance of the 
water circulating system. It is good practice to 
install a globe valve on the discharge side of the 
pump to regulate the water flow rate when the 
latter is critical. Too, where the piping is long 
and/or the quantity of water in the system is 
large, shut-off valves installed on the inlet and 
outlet of both the pump and the condenser will 
permit repairs to these pieces of equipment 
without the necessity of draining the tower. A 
drain connection should be installed at the 
lowest point in the piping and the piping should 
be pitched downward so as to assure complete 
drainage during winter shut down. 

The pump must always be located at some 
point below the level of the water in the tower 
basin in order to assure positive and continuous 
priming of the pump. When quiet operation is 
required, the pump may be isolated from the 
piping with short lengths of rubber hose. Auto- 
mobile radiator hose is suitable for this purpose. 

PROBLEMS 

1. A water piping system consists of 13S ft of 
2.5 in. nominal Type L smooth copper tube, 5 
standard elbows, and 2 globe valves (full open). 



If the flow rate through the pipe is 60 gpm, 
determine: 

(a) The total equivalent length of straight 
pipe. Ans. 297.5 equivalent ft 

(b) The total friction loss through the pipe is 
feet of water column. Ans. 9.28 ft H 2 

2. Rework Problem 1 using fairly rough pipe. 

Ans. 12.02 ft H 2 

3. A recirculating condenser water system con- 
sists of 100 ft of straight pipe, and 6 standard 
90° elbows. At the desired flow rate the pressure 
drop through the condenser is 7.5 psi and the 
pressure drop over the tower is 10 ft of water 
column. If 60 gpm are circulated through the 
system, determine: 

(a) The total equivalent length of pipe. 

Ans. 1 30 equivalent ft 

(b) The total head against which the pump 
must operate. Ans. 42.92 ft H 2 

4. From the manufacturer's rating curves, 
select a pump to fit the conditions of Problem 3. 

5. For a Refrigerant- 12 system, select a water 
regulating valve to meet the following con- 
ditions: 

(a) Desired condensing temperature range — 

90° to 105°. 
(6) Maximum entering water temperature — 

85° F. 

(c) Desired water quantity through condenser 
at maximum loading — 9 gpm. 

(d) Pressure available at city main during 
period of peak loading — 50 psi. 

(e) Pressure loss through condenser and water 
piping — 12 psi. 



16 

Refrigerants 



16-1. The Ideal Refrigerant. Generally speak- 
ing, a refrigerant is any body or substance which 
acts as a cooling agent by absorbing heat from 
another body or substance. With regard to the 
vapor-compression cycle, the refrigerant is the 
working fluid of the cycle which alternately 
vaporizes and condenses as it absorbs and gives 
off heat, respectively. To be suitable for use as a 
refrigerant in the vapor-compression cycle, a 
fluid should possess certain chemical, physical, 
and thermodynamic properties that make it both 
safe and economical to use. 

It should be recognized at the onset that there 
is no "ideal" refrigerant and that, because of the 
wide differences in the conditions and require- 
ments of the various applications, there is no one 
refrigerant that is universally suitable for all 
applications. Hence, a refrigerant approaches 
the "ideal" only to the extent that its properties 
meet the conditions and requirements of the 
application for which it is to be used. 

Table 16-1 lists a number of fluids having 
properties which render them suitable for use as 
refrigerants.* However, it will be shown 
presently that only a few of the more desirable 
ones are actually employed as such. Some, used 

* Since some of the fluorocarbon refrigerants, 
first introduced to the industry under the trade 
name "Freon," are now produced under several 
different trade marks, the ASRE, in order to avoid 
the confusion inherent in the use of either pro- 
prietary or chemical names, has adopted a number- 
ing system for the identification of the various 
refrigerants. Table 16-1 lists the ASRE number 
designation, along with the chemical name and 
formula for each of the compounds listed. 



extensively as refrigerants in the past, have been 
discarded as more suitable fluids were developed. 
Others, still in the development stage, show 
promise for the future. Tables 16-2 through 
16-6- list the thermodynamic properties of some 
of the refrigerants in common use at the present 
time. The use of these tables has already been, 
described in an earlier chapter. 
16-2. Safe Properties. Ordinarily, the safe 
properties of the refrigerant are the prime con- 
sideration in the selection of a refrigerant. It is 
for this reason that some fluids, which otherwise 
are highly desirable as refrigerants, find only 
limited use as such. The more prominent of 
these are ammonia and some of the straight 
hydrocarbons. 

To be suitable for use as a refrigerant, a fluid 
should be chemically inert to the extent that it is 
nonflammable, nonexplosive, and nontoxic both 
in the pure state and when mixed in any propor- 
tion with air. Too, the fluid should not react 
unfavorably with the lubricating oil or with any 
material normally used in the construction of 
refrigerating equipment. Nor should it react 
unfavorably with moisture which despite strin- 
gent precautions is usually present at least to 
some degree in all refrigerating systems. Fur- 
thermore, it is desirable that the fluid be of such a 
nature that it will not contaminate in any way 
foodstuff or other stored products in the event 
that a leak develops in the system. 
16-3. Toxicity. Since all fluids other than air 
are toxic in the sense that they will cause 
suffocation when in concentrations large enough 
to preclude sufficient oxygen to sustain life, 
toxicity is a relative term which becomes mean- 
ingful only when the degree of concentration 
and the time of exposure required to produce 
harmful effects are specified. 

The toxicity of most commonly used refrig- 
erants has been tested by National Fire Under- 
writers. As a result, the various refrigerants are 
separated into six groups according to their 
degree of toxicity, the groups being arranged in 
descending order (Column 1 of Table 16-7). 
Those falling into Group 1 are highly toxic and 
are capable of causing death or serious injury 
in relatively small concentrations and/or short 
exposure periods. On the other hand, those 
classified in Group 6 are only mildly toxic, being 
capable of causing harmful effects only in rela- 
tively large concentrations. Since injury from 



284 



REFRIGERANTS 285 



the latter group is caused more by oxygen 
deficiency than by any harmful effects of the 
fluids themselves, for all practical purposes the 
fluids in Group 6 are considered to be nontoxic. 
However, it should be pointed out that some 
refrigerants, although nontoxic when mixed with 
air in their normal state, are subject to decom- 
position when they come in contact with an open 
flame or an electrical heating element. The 
products of decomposition thus formed are 
highly toxic and capable of causing harmful 
effects in small concentrations and on short 
exposure. This is true of all the fluorocarbon 
refrigerants (see Column 3 of Table 16-7). 
16-4. Flammability and Explosiveness. 
With regard to flammability and explosiveness, 
most of the refrigerants in common use are 
entirely nonflammable and nonexplosive 
(Column 2 of Table 16-7). Notable exceptions 
to this are ammonia and the straight hydro- 
carbons. Ammonia is slightly flammable and 
explosive when mixed in rather exact propor- 
tions with air. However with reasonable 
precautions, the hazard involved in using 
ammonia as a refrigerant is negligible. 

Straight hydrocarbons, on the other hand, 
are highly flammable and explosive, and their 
use as refrigerants except in special applications 
and under the surveillance of experienced 
operating personnel is not usually permissible. 
Because of their excellent thermal properties, 
the straight hydrocarbons are frequently 
employed in ultra-low temperature applications. 
In such installations, the hazard incurred by 
their use is minimized by the fact that the 
equipment is constantly attended by operating 
personnel experienced in the use and handling 
of flammable and explosive materials. 

The "American Standard Safety Code for 
Mechanical Refrigeration" sets forth in detail 
the conditions and circumstances under which 
the various refrigerants can be safely used. 
Most local codes and ordinances governing 
the use of refrigerating equipment are based on 
this code, which is sponsored jointly by the 
ASRE and ASA. 

The degree of hazard incurred by the use of 
toxic refrigerants depends upon a number of 
factors, such as the quantity of refrigerant used 
with relation to the size of the space into which 
the refrigerant may leak, the type of occupancy, 
whether or not open flames are present, the 



odor of the refrigerant, and whether or not 
experienced personnel are on duty to attend the 
equipment. For example, a small quantity of 
even a highly toxic refrigerant presents little 
hazard when used in relatively large spaces in 
that it is not possible in the event of a leak for 
the concentration to reach a harmful level. 
Too, the danger inherent in the use of toxic 
refrigerants is somewhat tempered by the fact 
that toxic refrigerants (including decomposition 
products) have very noticeable odors which tend 
to serve as a warning of their presence. Hence, 
toxic refrigerants are usually a hazard only to 
infants and others who, by reason of infirmity or 
confinement, are unable to escape the fumes. 
At the present time, ammonia is the only toxic 
refrigerant that is used to any great extent, and 
its use is ordinarily limited to packing plants, 
ice plants, and large cold storage facilities where 
experienced personnel are usually on duty. 
16-5. Economic and Other Considerations. 
Naturally, from the viewpoint of economical 
operation, it is desirable that the refrigerant 
have physical and thermal characteristics which 
will result in the minimum power requirements 
per unit of refrigerating capacity, that is, a high 
coefficient of performance. For the most part, 
the properties of the refrigerant which influence 
the coefficient of performance are: (1) the 
latent heat of vaporization, (2) the specific 
volume of the vapor, (3) the compression ratio, 
and (4) the specific heat of the refrigerant in 
both the liquid and vapor states. 

Except in very small systems, a high latent 
heat value is desirable in that the weight of 
refrigerant circulated per unit of capacity is 
less. When a high latent heat value is accom- 
panied by a low specific volume in the vapor 
state, the efficiency and capacity of the com- 
pressor are greatly increased. This tends not 
only to decrease the power consumption but 
also to reduce the compressor displacement 
required, which permits the use of smaller, more 
compact equipment. However, in small systems, 
if the latent heat value of the refrigerant is too 
high, the amount of refrigerant circulated will 
be insufficient for accurate control of the liquid. 

A low specific heat for the liquid and a high 
specific heat for the vapor are desirable in that 
both tend to increase the refrigerating effect 
per pound, the former by increasing the sub- 
cooling effect and the latter by decreasing the 



286 PRINCIPLES OF REFRIGERATION 



Refrigerent-50 

Methane 

(CH4) -259°F 




Refrigerant-40 
Methyl Chloride 
(CH3CI) -11*F 



Refrigerant-30 
Methylene Chloride 
(CH2CI2) 104"F 



Refrigerant-20 
Chloroform 
(CHCI3) 142-F 





Refrigerant-21 
Dichloromonofluoromethane 
(CHCI2F) 48'F 



Refrigerant-22 
Monochlorodifluoromethane 
(CHCIF2) -41°F 



Refrigerant-11 
Trichloromonofluoromethane 
(CClaF) 75'F 



Refrigerant- 12 
Dichlorodifluoromethane 
(CCI2F2) -22°F 



Refrigerant- 13 
Monochlorotrifluoromethane 
(CCIF3) — 115*F 



Refrigerant-10 
Carbontetrach loride 
(CCI4) 169°F 




Refrigerant- 14 
Carbontetrafluoride 
(CF 4 ) -198°F 



Fig. 16-1. Methane series refrigerants. 



REFRIGERANTS 287 



superheating effect. When both are found in a 
single fluid, the efficiency of a liquid-suction 
heat exchanger is much improved. 

The effect of compression ratio on the work 
of compression and, consequently, on the 
coefficient of performance, has already been 
discussed in a previous chapter. Naturally, all 
other factors being equal, the refrigerant giving 
the lowest compression ratio is the most 
desirable. Low compression ratios result in 
low power consumption and high volumetric 
efficiency, the latter being more important in 
smaller systems since it permits the use of 
small compressors. 

Too, it is desirable that the pressure-tempera- 
ture relationship of the refrigerant is such 
that the pressure in the evaporator is always 
above atmospheric. In the event of a leak on the 
low pressure side of the system, if the pressure 
in the low side is below atmospheric, consider- 
able amounts of air and moisture may be drawn 
into the system, whereas if the vaporizing 
pressure is above atmospheric, the possibility of 
drawing in air and moisture in the event of a 
leak is minimized. 

Reasonably low condensing pressures under 
normal atmospheric conditions are also desirable 
in that they allow the use of lightweight 
materials in the construction of the condensing 
equipment, thereby reducing the size, weight, 
and cost of the equipment. 

Naturally, the critical temperature and 
pressure of the refrigerant must be above the 
maximum temperature and pressure which will 
be encountered in the system. Likewise, the 
freezing point of the refrigerant must be safely 
below the minimum temperature to be obtained 
in the cycle. These factors are particularly 
important in selecting a refrigerant for a low 
temperature application. 

In Table 16-8, a comparison is given of the 
performance of the various refrigerants at 
standard ton conditions (5° F evaporator and 
86° F condensing). Notice particularly that, 
with the exception of air, carbon dioxide, and 
ethane, the horsepower required per ton of 
refrigeration is very nearly the same for all the 
refrigerants listed. For this reason, efficiency 
and economy of operation are not usually 
deciding factors in the selection of the refrig- 
erant. More important are those properties 
which tend to reduce the size, weight, and 



initial cost of the refrigerating equipment and 
which permit automatic operation and a 
minimum of maintenance. 
16-6. Early Refrigerants. In earlier days, when 
mechanical refrigeration was limited to a 
few large applications, ammonia and carbon 
dioxide were practically the only refrigerants 
available. Later, with the development of 
small, automatic domestic and commercial 
units, refrigerants such as sulfur dioxide and 
methyl chloride came into use, along with 
methylene chloride, which was developed for 
use with centrifugal compressors. Methylene 
chloride and carbon dioxide, because of their 
safe properties, were extensively used in large 
air conditioning applications. 

With the exception of ammonia, all these 
refrigerants have fallen into disuse and are 
found only in some of the older installations, 
having been discarded in favor of the more 
suitable fluorocarbon refrigerants as the latter 
were developed. The fluorocarbons are practi- 
cally the only refrigerants in extensive use at the 
present time. Again, an exception to this is 
ammonia which, because of its excellent thermal 
properties, is still widely used in such instal- 
lations as ice plants, skating rinks, etc. A few 
other refrigerants also find limited use in special 
applications. 

16-7. Development of the Fluorocarbons. 
The search for a completely safe refrigerant 
with good thermal properties led to the develop- 
ment of the fluorocarbon refrigerants in the late 
1920's. The fluorocarbons (fluoronated hydro- 
carbons) are one group of a family of compounds 
known as the halocarbons (halogenated hydro- 
carbons). The halocarbon family of compounds 
are synthesized by replacing one or more of the 
hydrogen atoms in methane (CH*) or ethane 
(CgHg) molecules, both of which are pure 
hydrocarbons, with atoms of chlorine, fluorine, 
and/or bromine, the latter group comprising the 
halogen family. Halocarbons developed from 
the methane molecule are known as "methane 
series halocarbons." Likewise, those developed 
from the ethane molecule are referred to as 
"ethane series halocarbons." 

The composition of the methane series halo- 
carbons is shown in Fig. 16-1. Notice that the 
basic methane molecule consists of one atom of 
carbon (Q and four atoms of hydrogen (H). 
If the hydrogen atoms are replaced progressively 



288 PRINCIPLES OF REFRIGERATION 



Refrigerant- 170 

Ethane 

(CH3CH3) -127.5-F 





-113 

Trichlorotrifluoroethane 
(CCI2FCCIF2) 117.6'F 



114 

Dichlortetrafluofoethane 
(CClFjjCClFii) 38.4°F 



Fig. 16-2. Ethane series refrigerants. 



with chlorine (CI) atoms, the resulting com- 
pounds are methyl chloride (CH 3 C1), methylene 
chloride (CHaClg), chloroform (CHCI3), and 
carbontetrachloride (CC1 4 ), respectively, the 
last two being the base molecules for the more 
popular fiuorocarbons of the methane series. 

If the chlorine atoms in the carbontetra- 
chloride molecule are now replaced progressively 
with fluorine atoms, the resulting compounds 
are trichloromonofluoromethane (CC1 S F), 
dichlorodifluoromethane (CCl a F 2 ), mono- 
chlorotrifluoromethane (CCIF3), and carbon- 
tetrafiuoride (CF^, respectively. In the same 
order, the ASRE refrigerant standard number 
designations for these compounds are Refrig- 
erants-11, 12, 13, and 14, the last figure in the 
numbers being an indication of the number of 
fluorine atoms in the molecule. 

The molecular structure of Refrigerants-21 
and 22, which are also fiuorocarbons of the 
methane series, is shown in Fig. 16-1. Notice 
the presence of the hydrogen atom in each of 
these two compounds, an indication that they 
are derivatives of the chloroform molecule 
rather than the carbontetrachloride molecule. 

Figure 16-2 shows the molecular structure of 
Refrigerants-113 and 114, the only two fiuoro- 



carbons of the ethane series in common use. 
The presence of the two carbon atoms identifies 
the basic molecule as ethane, rather than 
methane, which has only one carbon atom. 
The individual characteristics of these and 
other refrigerants are discussed in the following 
sections. 

16-8. The Effect of Moisture. It is a well- 
established fact that moisture will combine in 
varying degrees with most of the commonly 
used refrigerants, causing the formation of 
highly corrosive compounds (usually acids) 
which will react with the lubricating oil and 
with other materials in the system, including 
metals. This chemical action often results in 
pitting and other damage to valves, seals, 
bearing journals, cylinder walls, and other 
polished surfaces. It may also cause deteriora- 
tion of the lubricating oil and the formation of 
metallic and other sludges which tend to clog 
valves and oil passages, score bearing surfaces, 
and otherwise reduce the life of the equipment. 
Moisture corrosion also contributes to com- 
pressor valve failure and, in hermetic motor- 
compressors, often causes breakdown of the 
motor winding insulation, which results in 
shorting or grounding of the motor. 



Although a completely moisture-free refrig- 
erating system is not possible, good refrig- 
erating practice demands that the moisture 
content of the system be maintained below the 
level which will produce harmful effects in the 
system. The minimum moisture level which 
will produce harmful effects in a refrigerating 
system is not clearly defined and will vary 
considerably, depending upon the nature of the 
refrigerant, the quality of the lubricating oil, and 
the operating temperatures of the system, par- 
ticularly the compressor discharge temperature. 

Moisture in a refrigerating system may exist 
as "free water" or it may be in solution with the 
refrigerant. When moisture is present in the 
system in the form of free water, it will freeze 
into ice in the refrigerant control and/or in the 
evaporator, provided that the temperature of 
the evaporator is maintained below the freezing 
point of the water. Naturally, the formation of 
ice in the refrigerant control orifice will prevent 
the flow of liquid refrigerant through that part 
and render the system inoperative until such 
time that the ice melts and flow through the 
control is restored. In such cases, refrigeration 
is usually intermittent as the flow of liquid is 
started and stopped by alternate melting and 
freezing of the ice in the control orifice. 

Since free water exists in the system only 
when the amount of moisture in the system 
exceeds the amount that the refrigerant can 
hold in solution, freeze-ups are nearly always 
an indication that the moisture content of the 
system is above the minimum level that will 
produce corrosion. On the other hand, the 
mere absence of freeze-ups cannot be taken to 
mean that the moisture content of the system 
is necessarily below the level which will cause 
corrosion, since corrosion can occur with some 
refrigerants at levels well below those which 
will result in free water. Too, it must be 
recognized that freeze-ups do not occur in air 
conditioning systems or in any other system 
where the evaporator temperature is above 
the freezing point of water. For this reason, 
high temperature systems are often more sub- 
ject to moisture corrosion than are systems 
operating at lower evaporator temperatures, 
since relatively large quantities of moisture can 
go unnoticed in such systems for relatively long 
periods of time. 

Since the ability of an individual refrigerant 



REFRIGERANTS 289 

to hold moisture in solution decreases as the 
temperature decreases, it follows that the 
moisture content in low temperature systems 
must be maintained at a very low level in order to 
avoid freeze-ups. Hence, moisture corrosion in 
low temperature systems is usually at a minimum. 

The various refrigerants differ greatly both as 
to the amount of moisture they will hold in 
solution and as to the effect that the moisture 
has upon them. For example, the straight 
hydrocarbons, such as propane, butane, ethane, 
etc., absorb little if any moisture. Therefore, 
any moisture contained in such systems will be 
in the form of free water and will make its 
presence known by freezing out in the refrig- 
erant control. Since this moisture must be 
removed immediately in order to keep the system 
operative, moisture corrosion will not usually 
be a problem when these refrigerants are used. 

Ammonia and sulfur dioxide, on the other 
hand, have an affinity for water and therefore 
are capable of absorbing moisture in such large 
quantities that free water is seldom found 
in systems employing these two refrigerants. 
However, the effects produced by the combina- 
tion of the water and the refrigerant are entirely 
different for the two refrigerants. 

In ammonia systems, the combination of 
water and ammonia produces aqua ammonia, a 
strong alkali, which attracts nonferrous metals, 
such as copper and brass, but has little if any 
effect on iron or steel or any other materials in 
the system. For this reason, ammonia systems 
can be operated successfully even when rela- 
tively large amounts of moisture are present in 
the system. 

In the case of sulfur dioxide, the moisture 
and sulfur dioxide combine to form sulfurous 
acid (H2SO3), which is highly corrosive. In 
view of the high solubility of water in S0 2 , the 
amount of acid formed can be quite large. 
Hence, corrosion in sulfur dioxide systems can 
be very heavy. 

The halocarbon refrigerants hydrolyze only 
slightly and therefore form only small amounts 
of acids or other corrosive compounds. As a 
general rule, corrosion will not occur in systems 
employing halocarbon refrigerants when the 
moisture content is maintained below the level 
which will cause freeze-ups, provided that high 
quality lubricating oils are used and that dis- 
charge temperatures are reasonably low. 



290 PRINCIPLES OF REFRIGERATION 



16-9. Refrigerant-Oil Relationship. With a 
few exceptions, the oil required for lubrication 
of the compressor is contained in the crankcase 
of the compressor where it is subject to contact 
with the refrigerant. Hence, as already stated, 
the refrigerant must be chemically and physically 
stable in the presence of oil, so that neither the 
refrigerant nor the oil is adversely affected by 
the relationship. 

Although some refrigerants, particularly sul- 
fur dioxide and the halocarbons, react with 
the lubricating oil to some extent, under normal 
operating conditions the reaction is usually 
slight and therefore of little consequence, 
provided that a high quality lubricating oil is 
used and that the system is relatively clean and 
dry. However, when contaminants, such as air 
and moisture, are present in the system in any 
appreciable amount, chemical reactions involv- 
ing the contaminants, the refrigerant, and the 
lubricating oil often occur which can result in 
decomposition of the oil, the formation of 
corrosive acids and sludges, copper plating, 
and/or serious corrosion of polished metal 
surfaces. High discharge temperatures greatly 
accelerate these processes, particularly oil de- 
composition, and often result in the formation of 
carbonaceous deposits on discharge valves and 
pistons and in the compressor head and dis- 
charge line. This condition is aggravated by the 
use of poorly refined lubricating oils containing 
a high percentage of unsaturated hydrocarbons, 
the latter being very unstable chemically. 

Because of the naturally high discharge 
temperature of Refrigerant-22 (see Table 16-8), 
breakdown of the lubricating oil, accompanied 
by motor burnouts, is a common problem with 
hermetic motor-compressor units employing 
this refrigerant, particularly when used in con- 
junction with air-cooled condensers and long 
suction lines. 

Copper plating of various compressor parts 
is often found in systems employing halocarbon 
refrigerants. The parts usually affected are the 
highly polished metal surfaces which generate 
heat, such as seals, pistons, cylinder walls, 
bearing surfaces, and valves. The exact cause of 
copper plating has not been definitely deter- 
mined, but considerable evidence does exist that 
moisture and poor quality lubricating oils 
are contributing factors. 

Because copper is never used with ammonia, 



copper plating is not found in ammonia systems. 
However, neither is it found in sulfur dioxide 
systems, although copper has been employed 
extensively with this refrigerant. 

In any event, regardless of the nature of 
and/or the cause of unfavorable reactions 
between the refrigerant and the lubricating oil, 
these disadvantages can be greatly minimized 
or eliminated by the use of high quality lubri- 
cating oils, having low "pour" and/or "floe" 
points (see Section 18-16), by maintaining the 
system relatively free of contaminants, such as 
air and moisture, and by designing the system so 
that discharge temperatures are reasonably low. 
16-10. Oil Miscibility. With regard to the 
refrigerant-oil relationship, one important 
characteristic which differs for the various 
refrigerants is oil miscibility, that is, the ability 
of the refrigerant to be dissolved into the oil and 
vice versa. 

With reference to oil miscibility, refrigerants 
may be divided into three groups: (1) those 
which are miscible with oil in all proportions 
under conditions found in the refrigerating 
system, (2) those which are miscible under 
conditions normally found in the condensing 
section, but separate from the oil under the 
conditions normally found in the evaporator 
section, and (3) those which are not miscible 
with oil at all (or only very slightly so) under 
conditions found in the system. 

As to whether or not oil miscibility is a 
desirable property in a refrigerant there is some 
disagreement. In any event, the fact of oil 
miscibility, or the lack of it, has little if any 
significance insofar as the selection of the 
refrigerant is concerned. However, since it 
greatly influences the design of the compressor 
and other system components, including the 
refrigerant piping, the degree of oil miscibility 
is an important refrigerant characteristic and 
therefore should be considered in some detail. 

With regard to the oil, one of the principal 
effects of an oil miscible refrigerant is to dilute 
the oil in the crankcase of the compressor, 
thereby lowering the viscosity (thinning) of the 
oil and reducing its lubricating qualities. To 
compensate for refrigerant dilution, the com- 
pressor lubricating oil used in conjunction with 
oil-miscible refrigerants should have a higher 
initial viscosity than that used for similar duty 
with nonmiscible refrigerants. 



REFRIGERANTS 291 



Viscosity may be defined as a measure of 
fluid friction or as a measure of the resistance 
that a fluid offers to flow. Hence, thin, low 
viscosity fluids will flow more readily than 
thicker, more viscous fluids. To provide 
adequate lubrication for the compressor, the 
viscosity of the lubricating oil must be main- 
tained within certain limits. If the viscosity of 
the oil is too low, the oil will not have sufficient 
body to form a protective film between the 
various rubbing surfaces and keep them 
separated. On the other hand, if the viscosity 
of the oil is too high, the oil will not have 
sufficient fluidity to penetrate between the 
rubbing surfaces, particularly where tolerances 
are close. In either case, lubrication of the 
compressor will not be adequate. 

Any oil circulating through the system with 
the refrigerant will have an adverse affect on 
the efficiency and capacity of the system, the 
principal reason being that the oil tends to 
adhere to and to form a film on the surface of 
the condenser and evaporator tubes, thereby 
lowering the heat transfer capacity of these two 
units. Since the oil becomes more viscous and 
tends to congeal as the temperature is reduced, 
the problem with oil is greatest in the evaporator 
and becomes more acute as the temperature of 
the evaporator is lowered. 

Since the only reason for the presence of oil 
in the refrigerating system is to lubricate the 
compressor, it is evident that the oil will best 
serve its function when confined to the com- 
pressor and not allowed to circulate with the 
refrigerant through other parts of the system. 
However, since, with few exceptions, the system 
refrigerant unavoidably comes into contact with 
the oil in the compressor, a certain amount of 
oil in the form of small particles will be en- 
trained in the refrigerant vapor and carried over 
through the discharge valves into the discharge 
line. If the oil is not removed from the vapor 
at this point, it will pass into the condenser and 
liquid receiver from where it will be carried 
to the evaporator by the liquid refrigerant. 
Obviously, in the interest of system efficiency 
and in order to maintain the oil in the crank- 
case at a constant level, some provision must be 
made for removing this oil from the system and 
returning it to the crankcase where it can per- 
form its lubricating function. 

The degree of difficulty experienced in bring- 



ing about the return of oil to the crankcase 
depends primarily on three factors: (1) the 
oil miscibility of the refrigerant, (2) the type of 
evaporator used, and (3) the evaporator tem- 
perature. 

When an oil-miscible refrigerant is employed, 
the problem of oil return is greatly simplified 
by the fact that the oil remains in solution with 
the refrigerant. This permits the oil to be 
carried along through the system by the 
refrigerant and, subsequently, to be returned to 
the crankcase through the suction line, provided 
that the evaporator and the refrigerant piping 
are properly designed. 

Unfortunately, when nonmiscible refriger- 
ants are used, once the oil passes into the 
condenser, the return of the oil to the crankcase 
is not so easily accomplished. The reason for 
this is that, except for a small amount of 
mechanical mixing, the refrigerant and the oil 
will remain separate, so that only a small 
portion of the oil is actually carried along with 
the refrigerant. For example, in the case of 
ammonia, which is lighter than oil, a large 
percentage of the oil will separate from the 
liquid ammonia and settle out at various low 
points in the system. For this reason, oil 
drains should be provided at the bottom of all 
receivers, evaporators, accumulators, and other 
vessels containing liquid ammonia, and pro- 
visions should be made for draining the oil 
from these points, either continuously or 
periodically, and returning it to the crankcase. 
This may be accomplished manually or auto- 
matically. 

When flooded-type evaporators are used, the 
refrigerant velocity will not usually be sufficient 
to permit the refrigerant vapor to entrain the 
oil and carry it over into the suction line and 
back to the crankcase. Hence, even with oil 
miscible refrigerants, where flooded-type evap- 
orators are employed, it is often necessary to 
make special provisions for oil return. The 
methods used to insure the continuous return 
of the oil from the evaporator to the crankcase 
in such cases is described in Chapter 19. 

Since the oil acts to lubricate the refrigerant 
flow control and other valves which may be in 
the system, the circulation of a small amount of 
oil with the refrigerant is not ordinarily objec- 
tionable. However, because of the adverse 
effect on system capacity, the amount of oil 



292 PRINCIPLES OF REFRIGERATION 



should be kept to a practical minimum. Too, 
since the oil in circulation comes initially from 
the compressor crankcase, an excessive amount 
in circulation may cause the oil level in the 
crankcase to fall below the minimum level 
required for adequate lubrication of the com- 
pressor parts. 

In order to minimize the circulation of oil, an 
oil separator or trap is sometimes installed 
in the discharge line between the compressor 
and the condenser (see Section 19-12). 

As a general rule, discharge line oil separators 
should be employed in any system where oil 
return is likely to be inadequate and/or where 
the amount of oil in circulation is apt to be 
excessive or to cause an undue loss in system 
capacity and efficiency. Specifically, discharge 
line oil separators are recommended for all 
systems employing nonmiscible refrigerants 
(or refrigerants which are not oil miscible at the 
evaporator conditions), not only because of the 
difficulty experienced in returning the oil from 
the evaporator to the crankcase but also because 
the presence of even small amounts of oil in the 
evaporators of such systems will usually cause 
considerable loss of evaporator efficiency and 
capacity. 

The same thing is usually true for systems 
employing miscible refrigerants when the 
evaporator temperature is below 0° F. Oil 
separators are recommended also for all 
systems using flooded evaporators, since oil 
return from this type of evaporator is apt to 
be inadequate because of low refrigerant 
velocities. 

Although oil separators are very effective in 
removing oil from the refrigerant vapor, they 
are not 100% efficient. Therefore, even though 
an oil separator is used, some means must still 
be provided for returning to the crankcase the 
small amount of oil which will always pass 
through the separator and find its way into 
other parts of the system. Too, since oil 
separators can often cause serious problems in 
the system if they are not properly installed, 
the use of oil separators should ordinarily be 
limited to those systems where the nature of the 
refrigerant or the particular design of the 
system requires their use. Oil separators are 
discussed in more detail in Chapter 19. 
16-11. Leak Detection. Leaks in a refrigerat- 
ing system may be either inward or outward, 



depending on whether the pressure in the 
system at the point of leakage is above or below 
atmospheric pressure. When the pressure in 
the system is above atmospheric at the point of 
leakage, the refrigerant will leak from the 
system to the outside. On the other hand, 
when the pressure in the system is below 
atmospheric, there is no leakage of refrigerant 
to the outside, but air and moisture will be 
drawn into the system. In either case, the 
system will usually become inoperative in a 
very short time. However, as a general rule, 
outward leaks are less serious than inward ones, 
usually requiring only that the leak be found and 
repaired and that the system be recharged with 
the proper amount of refrigerant. In the case of 
inward leaks, the air and moisture drawn into 
the system increase the discharge pressure and 
temperature and accelerates the rate of corro- 
sion. The presence of moisture in the system 
may also cause freeze-up of the refrigerant 
control. Furthermore, after the leak has been 
located and repaired, the system must be 
completely evacuated and dehydrated before it 
can be placed in operation. A refrigerant drier 
should also be installed in the system. 

The necessity of maintaining the system free 
of leaks demands some convenient means for 
checking a new system for leaks and for 
detecting leaks if and when they occur in 
systems already in operation. New systems 
should be checked for leaks under both vacuum 
and pressure. 

One method of leak detection universally 
used with all refrigerants employs a relatively 
viscous soap solution which is relatively free of 
bubbles. The soap solution is first applied to 
the pipe joint or other suspected area and then 
examined with the help of a strong light. The 
formation of bubbles in the soap solution 
indicates the presence of a leak. For adequate 
testing with a soap solution, the pressure in the 
system should be 50 psig or higher. 

The fact that sulfur and ammonia vapors 
produce a dense white smoke (ammonia 
sulfite) when they come into contact with one 
another provides a convenient means of 
checking for leaks in both sulfur dioxide and 
ammonia systems. To check for leaks in a 
sulfur dioxide system, a cloth swab saturated 
with stronger ammonia (approximately 28% 
available in any drug store) is held near, but 



not in contact with, all pipe joints and other 
suspected areas. A leak is indicated when the 
ammonia swab gives off a white smoke. 

Ammonia systems are checked in the same 
way except that a sulfur candle is substituted for 
the ammonia swab. Dampened phenophthalein 
paper, which turns red on contact with am- 
monia vapor, may also be used to detect 
ammonia leaks. 

A halide torch is often used to detect leaks in 
systems employing any of the halocarbon refrig- 
erants. The halide torch consists of a copper 
element which is heated by a flame. Air to 
support combustion is drawn in through a 
rubber tube, one end of which is attached to the 
torch. The free end of the tube is passed around 
all suspected areas. The presence of a halo- 
carbon vapor is indicated when the flame 
changes from its normal color to a bright green 
or purple. The halide torch should be used only 
in well-ventilated spaces. 

For carbon dioxide and the straight hydro- 
carbons, the only method of leak detection is 
the soap solution previously mentioned. 
16-12. Ammonia. Ammonia is the only refrig- 
erant outside of the fluorocarbon group that is 
being used to any great extent at the present 
time. Although ammonia is toxic and also 
somewhat flammable and explosive under cer- 
tain conditions, its excellent thermal properties 
make it an ideal refrigerant for ice plants, 
packing plants, skating rinks, large cold storage 
facilities, etc., where experienced operating 
personnel are usually on duty and where its 
toxic nature is of little consequence. 

Ammonia has the highest refrigerating effect 
per pound of any refrigerant. This, together 
with a moderately low specific volume in the 
vapor state, makes possible a high refrigerating 
capacity with a relatively small piston dis- 
placement. 

The boiling point of ammonia at standard 
atmospheric pressure is -28° F. The evapora- 
tor and condenser pressures at standard ton 
conditions of 5° F and 86° F are 19.6 psig and 
154.5 psig, respectively, which are moderate, so 
that lightweight materials can be used in the 
construction of the refrigerating equipment. 
However, the adiabatic discharge temperature 
is relatively high, being 210° F at standard ton 
conditions, which makes water cooling of the 
compressor head and cylinders desirable. Too, 



REFRIGERANTS 293 

high suction superheats should be avoided in 
ammonia systems. 

Although pure anhydrous ammonia is non- 
corrosive to all metals normally used in refrig- 
erating systems, in the presence of moisture, 
ammonia becomes corrosive to nonferrous 
metals, such as copper and brass. Obviously, 
these metals should never be used in ammonia 
systems. 

Ammonia is not oil miscible and therefore 
will not dilute the oil in the compressor crank- 
case. However, provisions must be made for 
the removal of oil from the evaporator and an 
oil separator should be used in the discharge line 
of all ammonia systems. 

Ammonia systems may be tested for leaks 
with sulfur candles, which give off a dense white 
smoke in the presence of ammonia vapor, or by 
applying a thick soap solution around the pipe 
joints, in which case a leak is indicated by the 
appearance of bubbles in the solution. 
16-13. Sulfur Dioxide. Sulfur dioxide (SO s ) 
is produced from the combustion of sulfur. It 
is highly toxic, but nonflammable and 
nonexplosive. In the 1920s and 1930s, sulfur 
dioxide was widely used in domestic refrigerators 
and in small commercial fixtures. Today, it is 
found only in a few of the older commercial 
units, having been replaced first by methyl 
chloride and later by the more desirable fluoro- 
carbon refrigerants. 

The boiling point of sulfur dioxide at atmos- 
pheric pressure is approximately 14° F. Satur- 
ation pressures at standard ton conditions of 
5°F and 86° F are 5.9 in. Hg and 51.8 psig, 
respectively. 

Sulfur dioxide is not oil miscible. However, 
unlike ammonia and carbon dioxide, liquid sul- 
fur dioxide is heavier than oil so that the oil 
floats on top of the refrigerant. Since this 
characteristic simplifies the problem of oil 
return, it accounts for the popularity enjoyed by 
sulfur dioxide in the past for small automatic 
equipment. 

Like most common refrigerants, sulfur dioxide 
in the pure state is noncorrosive to metals 
normally used in the refrigerating system. How- 
ever, it combines with moisture to form 
sulfurous acid (H 2 SOg) and sulfuric acid 
(HgSO^, both of which are highly corrosive. 
16-14. Carbon Dioxide. Carbon dioxide (COj) 
is one of the first refrigerants used in mechanical 



294 PRINCIPLES OF REFRIGERATION 



refrigerating systems. It is odorless, nontoxic, 
nonflammable, nonexplosive, and noncorrosive. 
Because of its safe properties, it has been widely 
used in the past for marine service and for air 
conditioning in hospitals, theaters, hotels, and 
in other places where safety is the prime con- 
sideration. Although a few of these older 
installations are still in service, at the present 
time the use of carbon dioxide as a refrigerant 
is limited for the most part to extremely low 
temperature applications, particularly in the 
production of solid C0 2 (dry ice). 

One of the chief disadvantages of carbon 
dioxide is its high operating pressures, which 
under standard ton conditions of 5° F and 86° F 
are 317.5 psig and 1031 psig, respectively. 
Naturally, this requires the use of extra heavy 
piping and equipment. However, because of the 
high vapor density of CO a , the volume of vapor 
handled by the compressor is only 0.96 cu ft per 
minute per ton at 5° F, so that compressor sizes 
are small. 

Another disadvantage of carbon dioxide is 
that the horsepower required per ton is approxi- 
mately twice that of any of the commonly used 
refrigerants. For carbon dioxide, the theoretical 
horsepower required per ton at standard con- 
ditions is 1.84, whereas for ammonia, the horse- 
power required per ton is only 0.989, the latter 
value being typical for most refrigerants. 

Since its boiling temperature at atmospheric 
pressure ( — 109.3° F) is below its freezing tem- 
perature (— 69.9° F) at this pressure, carbon 
dioxide cannot exist in the liquid state at atmos- 
pheric pressure nor at any pressure below its 
triple point pressure of 75.1 psia. At any 
pressure under 75.1 psia, solid carbon dioxide 
sublimes directly into the vapor state and there- 
fore below this pressure is found only in the 
solid and vapor states. Because of the low 
critical temperature of CO a (87.8° F), relatively 
low condensing temperatures are required for 
liquefaction. Carbon dioxide is nonmiscible in 
oil and therefore will not dilute the oil in the 
crankcase of the compressor. Like ammonia, 
it is lighter than oil. Hence, oil return problems 
are similar to those encountered in an ammonia 
system. 

Leak detection is by soap solution only. 
16-15. Methyl Chloride. Methyl chloride 
(CH 3 C1) is a halocarbon of the methane series. 
It has many of the properties desirable in a 



refrigerant, which accounts for its wide use in 
the past in both domestic and commercial appli- 
cations. Its boiling point at atmospheric 
pressure is —10.65° F. Evaporator and con- 
denser pressures at standard ton conditions are 
6.5 psig and 80 psig, respectively. 

Although methyl chloride is considered non- 
toxic, in large concentrations it has an anesthetic 
effect similar to that of chloroform, a compound 
to which it is closely related. Methyl chloride is 
moderately flammable and is explosive when 
mixed with air in concentrations between 8.1 
and 17.2% by volume. The hazard resulting 
from these properties is the principal reason for 
the discarding of methyl chloride in favor of the 
safer fluorocarbon refrigerants. 

Methyl chloride is corrosive to aluminum, 
zinc, and magnesium, and the compounds 
formed in combination with these materials are 
both flammable and explosive. Hence, these 
metals should not be used in methyl chloride 
systems. In the presence of moisture, methyl 
chloride forms a weak hydrochloric acid, which 
is corrosive to both ferrous and nonferrous 
metals. Too, since natural rubber and the 
synthetic, Neoprene, are dissolved by methyl 
chloride, neither is suitable gasket material for 
use in methyl chloride systems. 

Oil return in methyl chloride systems is 
simplified by the fact that methyl chloride is oil 
miscible. However, in selecting the compressor 
lubricating oil, crankcase dilution must be taken 
into account. 

Leaks in a methyl chloride system are found 
with the aid of a soap solution which is applied 
to the suspected joints. The presence of methyl 
chloride vapor may be detected with a halide 
leak detector. However, this method is not 
recommended because of the flammability of 
methyl chloride. 

16-16. Methylene Chloride (Carrene I). 
Methylene chloride (CH 2 C1 2 ), another halo- 
carbon of the methane series, has a boiling point 
of 103.5° F at atmospheric pressure, a charac- 
teristic which permits the refrigerant to be stored 
in sealed cans rather than in compressed gas 
cylinders. Under standard ton conditions, the 
evaporator and condenser pressures are both 
below atmospheric pressure, being 27.6 in. Hg 
and 9.5 in. Hg, respectively. Since the volume 
of the vapor handled per ton of refrigerating 
capacity is quite large (74.3 cu ft/min/ton at 



REFRIGERANTS 295 



5° F), centrifugal compressors, which are par- 
ticularly suited to handling large volumes of low 
pressure vapor, are required. 

Although it dissolves natural rubber, methy- 
lene chloride is noncorrosive even in the presence 
of moisture. It is also nontoxic and nonflam- 
mable. Because of its safe properties, it has been 
widely used in large air conditioning installations. 

The fact that methylene chloride is oil miscible 
is of little consequence, since in centrifugal com- 
pressors the oil and refrigerant do not ordinarily 
come in contact with one another. 

A halide torch or soap solution may be used 
to detect leaks. However, the pressure in the 
system must be built up above atmospheric in 
either case. 

16-17. Refrigerant-ll. Refrigerant- 1 1 (CC1 3 F) 
is a fluorocarbon of the methane series and 
has a boiling point at atmospheric pressure of 
74.7° F. Operating pressures at standard ton 
conditions are 24 in. Hg and 3.6 psig, respec- 
tively, which is very similar to those of methylene 
chloride. Although the theoretical horsepower 
required at standard ton conditions (0.927) is 
approximately the same as that for methylene 
chloride, the compressor displacement required 
at these conditions (36.32 cu ft/min/ton) is only 
approximately one-half that required for methy- 
lene chloride. 

Like other fluorocarbon refrigerants, Refrig- 
erant- 11 dissolves natural rubber. However, 
it is noncorrosive, nontoxic, and nonflammable. 
The low operating pressures and the relatively 
high compressor displacement required necessi- 
tate the use of a centrifugal compressor. 

Refrigerant-ll is used mainly in the air 
conditioning of small office buildings, factories, 
department stores, theaters, etc. A halide torch 
may be used for leak detection. 
16-18. Refrigerant- 1 2. Although its suprem- 
acy is being seriously challenged in some areas 
by Refrigerant-22, Refrigerant- 12 (CCl 8 Fg) is by 
far the most widely used refrigerant at the 
present time. It is a completely safe refrigerant 
in that it is nontoxic, nonflammable, and non- 
explosive. Furthermore, it is a highly stable 
compound which is difficult to break down even 
under extreme operating conditions. However, 
if brought into contact with an open flame or 
with an electrical heating element, Refrigerant- 
12 will decompose into products which are 
highly toxic (see Section 16-3). 



Along with its safe properties, the fact that 
Refrigerant- 12 condenses at moderate pressures 
under normal atmospheric conditions and has a 
boiling temperature of — 21 °F at atmospheric 
pressure makes it a suitable refrigerant for use 
in high, medium, and low temperature appli- 
cations and with all three types of compressors. 
When employed in conjunction with multistage 
centrifugal type compressors, Refrigerant-12 
has been used to cool brine to temperatures as 
low as -110° F. 

The fact that Refrigerant-12 is oil miscible 
under all operating conditions not only sim- 
plifies the problem of oil return but also tends 
to increase the efficiency and capacity of the 
system in that the solvant action of the refrig- 
erant maintains the evaporator and condenser 
tubes relatively free of oil films which otherwise 
would tend to reduce the heat transfer capacity 
of these two units. 

Although the refrigerating effect per pound for 
Refrigerant-12 is relatively small as compared to 
that of some of the other popular refrigerants, 
this is not necessarily a serious disadvantage. 
In fact, in small systems, the greater weight of 
Refrigerant-12 which must be circulated is a 
decided advantage in that it permits closer 
control of the liquid. In larger systems, the 
disadvantage of the low latent heat value is 
offset somewhat by a high vapor density, so that 
the compressor displacement required per ton of 
refrigeration is not much greater than that 
required for Refrigerants-22, 500, and 717. The 
horsepower required per ton of capacity com- 
pares favorably with that required for other 
commonly used refrigerants. 

A halide torch is used for leak detection. 
16-19. Refrigerant- 1 3. Refrigerant-13(CC1F,) 
was developed for and is being used in ultra-low 
temperature applications, usually in the low 
stage of a two or three stage cascade system. It 
is also being used to replace Refrigerant-22 in 
some low temperature applications. 

The boiling temperature of Refrigerant- 13 is 
— 144.5° F at atmospheric pressure. Evaporator 
temperatures down to — 150°F are practical. 
The critical temperature is 83.9° F. Since con- 
densing pressures and the compressor displace- 
ment required are both moderate, Refrigerant-1 3 
is suitable for use with all three types of 
compressors. 

Refrigerant-13 is a safe refrigerant. It is not 



296 PRINCIPLES OF REFRIGERATION 



miscible with oil. A halide torch may be used 
for leak detection. 

16-20. Refrigerant-22. Refrigerant-22 
(CHClFj) has a boiling point at atmospheric 
pressure of —41.4° F. Developed primarily as a 
low temperature refrigerant, it is used exten- 
sively in domestic and farm freezers and in 
commercial and industrial low temperature 
systems down to evaporator temperatures as 
low as — 125°F. It also finds wide use in 
packaged air conditioners, where, because of 
space limitations, the relatively small compressor 
displacement required is a decided advantage. 

Both the operating pressures and the adiabatic 
discharge temperature are higher for Refrig- 
erant-22 than for Refrigerant-12. Horsepower 
requirements are approximately the same. 

Because of the high discharge temperatures 
experienced with Refrigerant-22, suction super- 
heat should be kept to a minimum, particularly 
where hermetic motor-compressors are em- 
ployed. In low temperature applications, where 
compression ratios are likely to be high, water 
cooling of the compressor head and cylinders is 
recommended in order to avoid overheating of 
the compressor. Air-cooled condensers used 
with Refrigerant-22 should be generously sized. 

Although miscible with oil at temperatures 
found in the condensing section, Refrigerant-22 
will often separate from the oil in the evaporator. 
The exact temperature at which separation 
occurs varies considerably with the type of oil 
and the amount of oil mixed with the refrigerant. 
However, no difficulty is usually experienced 
with oil return from the evaporator when a 
properly designed serpentine evaporator is used 
and when the suction piping is properly designed. 
When flooded evaporators are employed, oil 
separators should be used and special pro- 
visions should be made to insure the return of 
oil from the evaporator. Oil separators should 
always be used on low temperature applica- 
tions. 

The principal advantage of Refrigerant-22 
over Refrigerant-12 is the smaller compressor 
displacement required, being approximately 
60% of that required for Refrigerant-12. Hence, 
for a given compressor displacement, the refrig- 
erating capacity is approximately 60% greater 
with Refrigerant-22 man with Refrigerant-12. 
Too, refrigerant pipe sizes are usually smaller 
for Refrigerant-22 than for Refrigerant-12. For 



evaporator temperatures between —20 and 
—40° F, still another advantage added to 
Refrigerant-22 is that the evaporator pressures 
for Refrigerant-22 at these temperatures are 
above atmospheric, whereas for Refrigerant-12 
the evaporator pressures will be below atmos- 
pheric. However, all this should not be taken to 
mean that Refrigerant-22 is superior to Refrig- 
erant-12 in all applications. As a matter of fact, 
except in those applications where space limita- 
tions necessitate the use of the smallest possible 
equipment and/or where the evaporator tem- 
perature is between -20° F and -40° F, 
Refrigerant-12, because of its lower discharge 
temperatures and greater miscibility with oil, is 
probably the more desirable of the two 
refrigerants. 

The ability of Refrigerant-22 to absorb mois- 
ture is considerably greater than that of Refrig- 
erant-12 and therefore less trouble is experi- 
enced with freeze-ups in Refrigerant-22 systems. 
Although some consider this to be an advantage, 
the advantage gained is questionable, since any 
amount of moisture in a refrigerating system is 
undesirable. 

Being a fluorocarbon, Refrigerant-22 is a safe 
refrigerant. A halide torch may be used for leak 
detection. 

16-21. Refrigerant-1 13. Refrigerant-113 
(CCljjFCClFa) boils at 117.6° F under atmos- 
pheric pressure. Operating pressures at standard 
ton conditions are 27.9 in. Hg and 13.9 in. Hg, 
respectively. Although the compressor displace- 
ment per ton is somewhat high ( 1 00.76 cu ft/min/ 
ton at standard ton conditions), the horsepower 
required per ton compares favorably with other 
common refrigerants. The low operating 
pressures and the large displacement required 
necessitate the use of a centrifugal type 
compressor. 

Although used mainly in comfort air con- 
ditioning applications, it is also employed in 
industrial process water and brine chilling down 
to 0° F. 

Refrigerant-1 13 is a safe refrigerant. A halide 
torch may be used for leak detection. 
16-22. Refrigerant-1 14. Refrigerant- 114 
(CClgCClFj) has a boiling point of 38.4° F under 
atmospheric pressure. Evaporating and con- 
densing pressures at standard ton conditions are 
16.1 in. Hg and 22 psig, respectively. The com- 
pressor displacement required is relatively low 



REFRIGERANTS 297 



for a low pressure refrigerant (19.59 cu ft/min/ 
ton at standard conditions) and the horsepower 
required compares favorably with that required 
by other common refrigerants. 

Refrigerant- 1 14 is used with centrifugal com- 
pressors in large commercial and industrial air 
conditioning installations and for industrial 
process water chilling down to —70° F. It is 
also used with vane-type rotary compressors in 
domestic refrigerators and in small drinking 
water coolers. 

Like Refrigerant-22, Refrigerant- 1 14 is oil 
miscible under conditions found in the con- 
densing section, but separates from oil in the 
evaporator. However, because of the type of 
equipment used with Refrigerant- 1 14 and the 
conditions under which it is used, oil return is 
not usually a problem. 

Refrigerant-1 14 is a safe refrigerant. A halide 
torch may be used for leak detection. 
16-23. Straight Hydrocarbons. The straight 
hydrocarbons are a group of fluids composed in 
various proportions of the two elements hydro- 
gen and carbon. Those having significance as 
refrigerants are methane, ethane, butane, pro- 
pane, ethylene, and isobutane. All are extremely 
flammable and explosive. Too, since all act as 
anesthetics in varying degrees, they are con- 
sidered mildly toxic. Although none of these 
compounds will absorb moisture to any appre- 
ciable extent, all are extremely miscible with oil 
under all conditions. 

Although a few of the straight hydrocarbons 
(butane, propane, and isobutane) have been 
used in small quantities for domestic refrigera- 
tion, their use is ordinarily limited to special 
applications where an experienced attendent is 
on duty. Ethane, methane, and ethylene are 
employed to some extent in ultra-low tempera- 
ture applications, usually in the lower stage of 
two and three stage cascade systems. However, 
even in these applications, it is likely that they 
will be replaced in the future by Refrigerants- 13 
and 14, the latter being used only in pilot plants 
at the present time. 

Leak detection is by soap solution only. 
I6-Z4. Refrigerant-500. Refrigerant-500, 
commonly known as Carrene 7*, is an azeo- 
tropic mixturet of Refrigerant- 12 (73.8% by 

* A proprietary refrigerant of the Carrier Corpor- 
ation. 

f An azeotropic mixture is a mixture of two or 



weight) and Refrigerant- 152a (26.2%). It has a 
boiling point at atmospheric pressure of —28° F. 
Evaporator and condenser pressures at standard 
ton conditions are 16.4 psig and 113.4 psig, 
respectively. Although the horsepower require- 
ments of Refrigerant-500 are approximately the 
same as those for Refrigerants- 12 and 22, the 
compressor displacement required is greater 
than that required for Refrigerant-22, but some- 
what less than that required for Refrigerant- 12. 

The principal advantage of Refrigerant-500 
lies in the fact that its substitution for Refrig- 
erant- 12 results in an increase in compressor 
capacity of approximately 18%. This makes it 
possible to use the same direct connected com- 
pressor (as in a hermetic motor-compressor 
unit) on either 50 or 60 cycle power with little or 
no change in the refrigerating capacity or in the 
power requirements. 

It will be shown in Chapter 21 that the speed 
of an alternating current motor varies in direct 
proportion to the cycle frequency. Therefore, 
an electric motor operating on 50 cycle power 
will have only five-sixths of the speed it has when 
operating on 60 cycle power. For this reason, 
the displacement of a direct connected com- 
pressor is reduced approximately 18% when a 
change is made from 60 to 50 cycle power. 
Since the increase in capacity per unit of dis- 
placement accruing from the substitution of 
Refrigerant-500 for Refrigerant- 12 is almost 
exactly equal to the loss of displacement suffered 
when changing for 60 to 50 cycle power, the 
same motor-compressor assembly is made 
suitable for use with both frequencies by the 
simple expedient of changing refrigerants. 
16-25. Refrigerant Drying Agents. Refrig- 
erant drying agents, called desiccants are fre- 
quently employed in refrigerating systems to 
remove moisture from the refrigerant. Some 
of the most commonly used desiccants are silica 
gel (silicon dioxide), activated alumina (alu- 
minum oxide), and Drierite (anhydrous calcium 
sulfate). Silica gel and activated alumina are 
adsorption-type desiccants and are available in 
granular form. Drierite is an absorption type 
desiccant and is available in granular form and 
in cast sticks. 

more liquids, which, when mixed in precise pro- 
portions, form a compound having a boiling 
temperature which is independent of the boiling 
temperatures of the individual liquids. 



17 

Refrigerant Flow 
Controls 



17-1. Types and Function. There are six 
basic types of refrigerant flow controls: (1) the 
hand expansion valve, (2) the automatic expan- 
sion valve, (3) the thermostatic expansion valve, 
(4) the capillary tube, (5) the low pressure float, 
and (6) the high pressure float. 

Regardless of type, the function of any refrig- 
erant flow control is twofold: (1) to meter the 
liquid refrigerant from the liquid line into the 
evaporator at a rate commensurate with the rate 
at which vaporization of the liquid is occurring 
in the latter unit, and (2) to maintain a pressure 
differential between the high and low pressure 
sides of the system in order to permit the refrig- 
erant to vaporize under the desired low pres- 
sure in the evaporator while at the same time 
condensing at a high pressure in the condenser. 
17-2. Hand Expansion Valves. Hand ex- 
pansion valves are hand-operated needle 
valves (Fig. 17-1). The rate of liquid flow 
through the valve depends on the pressure 
differential across the valve orifice and on the 
degree of valve opening, the latter being manu- 
ally adjustable. Assuming that the pressure 
differential across the valve remains the same, 
the flow rate through a hand expansion valve 
will remain constant at all times without regard 
for either the evaporator pressure or the 
evaporator loading. 

The principal disadvantage of the hand 
expansion valve is that it is unresponsive to 
changes in the system load and therefore must be 
manually readjusted each time the load on the 
system changes in order to prevent either 



starving or overfeeding of the evaporator, 
depending upon the direction of the load shift. 
Too, the valve must be opened and closed 
manually each time the compressor is cycled on 
and off. 

Obviously the hand expansion valve is suitable 
for use only on large systems where an operator 
is on duty and where the load on the system is 
relatively constant. When automatic control is 
desired and/or when the system is subject to 
frequent load fluctuations, some other type of 
refrigerant flow control is required. 

At the present time, the principal use of the 
hand expansion valve is as an auxiliary refrig- 
erant control installed in a by-pass line (Fig. 
17-29). It is also frequently used to control the 
flow rate through oil bleeder lines (Fig. 19-12). 
17-3. Automatic Expansion Valves. A 
schematic diagram of an automatic expansion 
valve is shown in Fig. 17-2. The valve consists 
mainly of a needle and seat, a pressure bellows 
or diaphragm, and a spring, the tension of the 
latter being variable by means of an adjusting 
screw. A screen or strainer is usually installed 
at the liquid inlet of the valve in order to pre- 
vent the entrance of foreign materials which may 
cause stoppage of the valve. The construction 
of a typical automatic expansion valve is shown 
in Fig. 17-3. 

The automatic expansion valve functions to 
maintain a constant pressure in the evaporator 
by flooding more or less of the evaporator sur- 
face in response to changes in the evaporator 
load. The constant pressure characteristic of the 
valve results from the interaction of two opposing 
forces: (1) the evaporator pressure and (2) the 
spring pressure. The evaporator pressure, 
exerted on one side of the bellows or diaphragm, 
acts to move the valve in a closing direction, 
whereas the spring pressure, acting on the 
opposite side of the bellows or diaphragm, acts 
to move the valve in an opening direction. When 
the compressor is running, the valve functions 
to maintain the evaporator pressure in equi- 
librium with the spring pressure. 

As the name implies, the operation of the 
valve is automatic and, once the tension of the 
spring is adjusted for the desired evaporator 
pressure, the valve will operate automatically 
to regulate the flow of liquid refrigerant into the 
evaporator so that the desired evaporator 
pressure is maintained, regardless of evaporator 



298 



REFRIGERANT FLOW CONTROLS 299 



loading. For example, assume that the tension 
of the spring is adjusted to maintain a constant 
pressure in the evaporator of lOpsig. There- 
after, any time the evaporator pressure tends to 
fall below lOpsig, the spring pressure will 
exceed the evaporator pressure causing the valve 
to move in the opening direction, thereby 
increasing the flow of liquid to the evaporator 
and flooding more of the evaporator surface. 
As more of the evaporator surface becomes 
effective, the rate of vaporization increases and 



i" Flare 




0.078" Orifice 
f Flare 



Fig. 17-1. Small capacity hand-expansion valve. 
(Courtesy Mueller Brass Company.) 



the evaporator pressure rises until equilibrium 
is established with the spring pressure. Should 
the evaporator pressure tend to rise above the 
desired 10 psig, it will immediately override the 
pressure of the spring and cause the valve to 
move in the closing direction, thereby throttling 
the flow of liquid into the evaporator and 
reducing the amount of effective evaporator 
surface. Naturally, this decreases the rate of 
vaporization and lowers the evaporator pressure 
until equilibrium is again established with the 
spring pressure. 



Adjusting 
screw ~ 



Bellows or 
diaphragm"" ?"^. 

Needle 



Strainer 




Evaporator pressure 



Fig. 17-2. Schematic diagram of automatic expansion 
valve. 



It is important to notice that the operating 
characteristics of the automatic expansion valve 
are such that the valve will close off tightly when 
the compressor cycles off and remain closed 
until the compressor cycles on again. As pre- 
viously described, vaporization continues in the 
evaporator for a short time after the com- 
pressor cycles off and, since the resulting vapor 
is not removed by the compressor, the pressure 
in the evaporator rises. Hence, during the off 
cycle, the evaporator pressure will always exceed 
the spring pressure and the valve will be tightly 
closed. When the compressor cycles on, the 
evaporator pressure will be immediately reduced 
below the spring pressure, at which time the 
valve will open and admit sufficient liquid to the 
evaporator to establish operating equilibrium 
between the evaporator and spring pressures. 




Fig. 17-3. Typical automatic expansion valve. (Cour- 
tesy Controls Company of America.) 



300 PRINCIPLES OF REFRIGERATION 



Automatic 

- expansion 

valve 



Liquid to 
here 



Liquid 

from 

receiver 




(a) 



Vapor! 
compressor 



Automatic 

-expansion 

valve 




receiver 



Liquid to 
here 



(b) 



Vapor 1 
compressor 



Fig. 17-4. Operating characteristics of the auto- 
matic expansion valve under varying load conditions. 
(a) Heavy load conditions, (b) Minimum load 
conditions. 



The chief disadvantage of the automatic 
expansion valve is its relatively poor efficiency 
as compared to that of other refrigerant flow 
controls. In view of the evaporator-compressor 
relationship, it is evident that maintaining a 
constant pressure in the evaporator requires that 
the rate of vaporization in the evaporator be 
kept constant. To accomplish this necessitates 
severe throttling of the liquid in order to limit 
the amount of effective evaporator surface when 
the load on the evaporator is heavy and the heat 
transfer capacity per unit of evaporator surface 
is high (Fig. 17-4a). As the load on the evapora- 
tor decreases and the heat transfer capacity per 
unit of evaporator surface is reduced, more and 
more of the evaporator surface must be flooded 
with liquid if a constant rate of vaporization is 
to be maintained (Fig. 17-46). As a matter of 



fact, if the load on the evaporator is permitted to 
fall below a certain level, the automatic expan- 
sion valve, in an attempt to keep the evaporator 
pressure up, will overfeed the evaporator to the 
extent that liquid will enter the suction line and 
be carried to the compressor where it may cause 
serious damage. However, in a properly 
designed system, overfeeding is not likely to 
occur, since the thermostat will usually cycle the 
compressor off before the space or product 
temperature is reduced to a level such that the 
load on the evaporator will fall below the 
critical point. 

Obviously, since it permits only a small 
portion of the evaporator to be filled with liquid 
during periods when the load on the system is 
heavy, the constant pressure characteristic of the 
automatic expansion valve severely limits the 
capacity and efficiency of the refrigerating 
system at a time when high capacity and high 
efficiency are most desired. Too, because the 
evaporator pressure is maintained constant 
throughout the entire running cycle of the com- 
pressor, the valve must be adjusted for a 
pressure corresponding to the lowest evaporator 
temperature required during the entire running 
cycle (see Fig. 17-5). This results in a consider- 
able loss in compressor capacity and efficiency, 
since advantage cannot be taken of the higher 
suction temperatures which would ordinarily 
exist with a full-flooded evaporator during the 
early part of the running cycle. 

Another disadvantage of the automatic ex- 
pansion valve, which can also be attributed to 
its constant pressure characteristic, is that it 
cannot be used in conjunction with a low 
pressure motor control, since proper operation 
of the latter part depends on a rather substantial 
change in the evaporator pressure during the 
running cycle, a condition which obviously 



30 
&|15 



> (A 
LU 0) 



o. 5 «-0ff ->fr -0n->j.^0ff 



^ 




0n-»f«— Off 



5 10 15 20 25 30 35 40 45 50 55 60 
Time in minutes 

Fig. 17-5. Operating characteristics of the automatic 
expansion valve. 



REFRIGERANT FLOW CONTROLS 301 



cannot be met when an automatic expansion 
valve is used as the refrigerant flow control. 

In view of its poor efficiency under heavy load 
conditions, the automatic expansion valve is 
best applied only to small equipment having 
relatively constant loads, such as domestic refrig- 
erators and freezers and small, retail ice-cream 
storage cabinets. However, even in these appli- 
cations the automatic expansion valve is seldom 
used at the present time, having given way to 
other types of refrigerant flow controls which 
are more efficient and sometimes lower in 
cost. 

Some automatic expansion valves are now 
being employed as "condenser by-pass valves." 



evaporator pressure. In this respect, the con- 
denser by-pass serves the same function as the 
cylinder by-pass type of compressor capacity 
control. * However, unlike the cylinder by-pass, 
the condenser by-pass does not unload the 
compressor in any way. Hence, with the con- 
denser by-pass, there is no reduction in the work 
of compression or in the power requirements of 
the compressor. For this reason, the condenser 
by-pass is not generally recommended as a 
means of controlling the capacity of the 
compressor. 

Care should be taken to connect the by-pass 
line to the condenser at a point low enough on 
the condenser to insure that slightly "wet" 




C 



Fig. 17-6. Automatic expansion 
valve employed as condenser 
by-pass valve. 



^> r 




Thermostatic 
• ^^"""expansion valve 



J 



Automatic expansion valve 

adjusted for minimum desired 

evaporator pressure 



_ Condenser 
"by-pass line 



r\ 



c 



c 



") 



As such thay are installed in a by-pass line 
between the condenser and the suction line (Fig. 
17-6) where they serve to regulate the flow of hot 
gas which is by-passed from the condenser 
directly into the suction line in order to prevent 
the evaporator pressure from dropping below a 
predetermined desired minimum. In such cases, 
the valve is set for the minimum desired eva- 
porator pressure. As long as the pressure in the 
evaporator remains above the desired minimum, 
the valve will remain closed and no gas is by- 
passed from the condenser into the suction line. 
However, any time the evaporator tends to fall 
below the desired minimum, the by-pass valve 
opens and permits hot gas from the