• TABLE OF CONTENTS
HIDE
 Front Cover
 Front Matter
 Title Page
 Foreword
 Acknowledgement
 Table of Contents
 Introduction
 Thermodynamic principles
 Heat sources
 Performance of indirect expansion...
 The performance of direct expansion...
 Internal conbustion engine power...
 The economics of reverse cycle...
 Novel arrangements of equipment...
 Heat pump installations
 Conclusion
 Glossary of abbreviations
 Bibliography














Group Title: Bulletin
Title: Reverse cycle refrigeration for heating in the South
CITATION THUMBNAILS PAGE IMAGE ZOOMABLE
Full Citation
STANDARD VIEW MARC VIEW
Permanent Link: http://ufdc.ufl.edu/UF00003196/00001
 Material Information
Title: Reverse cycle refrigeration for heating in the South
Series Title: Bulletin
Physical Description: 48 p. : ill., map ; 23 cm.
Language: English
Creator: Goethe, S. P ( Sam Paul ), 1914-
University of Florida -- Engineering and Industrial Experiment Station
Publisher: Florida Engineering and Industrial Experiment Station, College of Engineering, University of Florida
Place of Publication: Gainesville
Publication Date: 1947
 Subjects
Subject: Refrigeration and refrigerating machinery   ( lcsh )
Heating -- Equipment and supplies -- Southern States   ( lcsh )
Genre: government publication (state, provincial, terriorial, dependent)   ( marcgt )
bibliography   ( marcgt )
non-fiction   ( marcgt )
 Notes
Bibliography: Includes bibliographical references (p. 48).
Statement of Responsibility: by S.P. Goethe.
General Note: "June, 1947."
 Record Information
Bibliographic ID: UF00003196
Volume ID: VID00001
Source Institution: University of Florida
Holding Location: University of Florida
Rights Management: All rights reserved by the source institution and holding location.
Resource Identifier: ltqf - AAA4329
ltuf - AJL4525
oclc - 27180413
alephbibnum - 001790861

Table of Contents
    Front Cover
        Front Cover
    Front Matter
        Front Matter
    Title Page
        Page i
        Page ii
    Foreword
        Page iii
    Acknowledgement
        Page iv
    Table of Contents
        Page v
    Introduction
        Page 1
        Page 2
        Page 3
    Thermodynamic principles
        Page 4
        Page 5
        Page 6
        Page 7
        Page 8
        Page 9
        Page 10
    Heat sources
        Page 11
        Page 12
        Page 13
        Page 14
        Page 15
        Page 16
    Performance of indirect expansion reverse cycle refrigeration heating system
        Page 17
        Page 18
        Page 19
        Page 20
        Page 21
    The performance of direct expansion reverse cycle refrigeration heating system
        Page 22
        Page 23
    Internal conbustion engine power for reverse cycle refrigeration
        Page 24
        Page 25
        Page 26
        Page 27
        Page 28
    The economics of reverse cycle heating
        Page 29
        Page 30
        Page 31
        Page 32
        Page 33
        Page 34
    Novel arrangements of equipment for reverse cycle refrigeration heating
        Page 35
        Page 36
        Page 37
    Heat pump installations
        Page 38
        Page 39
        Page 40
        Page 41
        Page 42
    Conclusion
        Page 43
        Page 44
        Page 45
        Page 46
    Glossary of abbreviations
        Page 47
    Bibliography
        Page 48
        Page 49
Full Text














Reverse Cycle Refrigeration

for Heating in the South

Dr
S. P. GOETHE
Research Engineer
Mechanical Engineering Section


Bulletin No. 14


THE FLORIDA ENGINEERING AND INDUSTRIAL EXPERIMENT STATION
College of Enineering University of Florida Gainesville


June, I947





The Florida Engineering and Industrial Experiment Station
The Engineering Experiment Station was first approved by the
Board of Control at its meeting on May 18, 1929. Funds for the Florida
Engineering and Industrial Experiment Station were appropriated by
the Legislature of the State of Florida in 1941. The Station is a Divi-
sion of the College of Engineering of the University of Florida under
the supervision of the State Board of Control of Florida. The functions
of the Engineering and Industrial Experiment Station are:
a) To develop the industries of Florida by organizing and pro-
moting research in those fields of engineering, and the related sciences,
bearing on the industrial welfare of the State.
b) To survey and evaluate the natural resources of the State that
may be susceptible to sound development.
c) To contract with governmental bodies, technical societies, as-
sociations, or industrial organizations in aiding them to solve their
technical problems. Provision is made for these organizations to avail
themselves of the facilities of the Engineering and Industrial Experi-
ment Station on a co-operative financial basis. It is the basic philosophy
of the Station that the industrial progress of Florida can best be
furthered by carrying on research in those fields in which Florida, by
virtue of its location, climate, and raw materials, has natural
advantages.
d) To publish and disseminate information on the results of ex-
perimental and research projects. Two series of pamphlets are issued:
Bulletins covering the results of research and investigations by staff
members; and Technical Papers, reprinting papers or reports by staff
members which have been published elsewhere.
For copies of Bulletins, Technical Papers or information on how
the Station can be of service, address:
The Florida Engineering and Industrial Experiment Station
College of Engineering
University of Florida
Gainesville, Florida
JOSEPH WEIL, Director








REVERSE CYCLE REFRIGERATION

FOR HEATING IN THE SOUTH


BY
S. P. GOETHE
Research Engineer
Mechanical Engineering Section


THE FLORIDA ENGINEERING AND INDUSTRIAL EXPERIMENT STATION
COLLEGE OF ENGINEERING UNIVERSITY OF FLORIDA 0 GAINESVILLE


BULLETIN NO. 14-JUNE, 1947















































Permission is given to reproduce or quote any portion of this
publication providing a credit line is given acknowledging the
source of the information.












FOREWORD
Florida's agreeable, year-round climate has been a boon not
only to the development of its many industries, but to the comfort
and economy of the citizens contributing to those pursuits. Indica-
tions have appeared which tend to show that the favorable climatic
conditions are conductive to increasing the productivity of per-
sonnel in industry. Yet there are short seasons during the year when
abrupt temperature changes make the availability of artificial heat-
ing and cooling equipment a necessity of prime importance.
In this project the research engineer points out that the dual
problem of heat and refrigeration can be efficiently and economic-
ally solved by use of one set of centrally located facilities. With
reverse cycle of refrigeration for heating, lie eliminates flues, smoke
pollution, fire hazards, and fuel tanks. lie also shows that cooling
can be had in one area of a building while heat is applied to
another.
It is hoped that the publishing of this work will enable em-
ployers in industry to study these possibilities with respect to
creating more healthful working conditions and maintaining the
general efficiency of their employees at all times.
It is especially desirable that interested citizens read these
pages critically for information which will aid them in economically
installing equipment for permitting more refreshing and pleasing
conditions domestically for their families.
JosEPH WVEIL, Director.














ACKNOWLEDGEMENTS
The author wishes to acknowledge the helpful criticism of
Professor N. C. Ebaugh, Head of the Mechanical Engineering De-
partment, College of Engineering, University of Florida, which
has been sincerely appreciated. Primary research on this subject
was carried on in the Department of Mechanical Engineering of
the University of Florida as a partial requirement for a Master of
Science in Engineering degree from 1936 to 1938, and the coopera-
tion of all members of the department in preparation of this
material is gratefully acknowledged.
Appreciation is expressed to the American Society of Refriger-
ating Engineers for a research grant which partially defrayed the
expenses incurred in the preparation of this material for publication.
Acknowledgment of engineering performance data on air con-
ditioning equipment is made to the Westinghouse Electric &
Manufacturing Company.
The author also wishes to acknowledge the aid of Mr. E. N.
Kemler, Head of the Engineering Research Division, Southern
Research Institute in making available certain data which was
gathered by their organization.









TABLE OF CONTENTS

Acknowledgements ..----. ......................... .. ------. -- iv
Abstract .....-.. .......---.....----- .... -. ------.-...- -.. -------- 1
Introduction .. ---.....- -- ... ..............--- --.--... ..... 1
Thermodynamic Principles .................. .......---- ..... .... 4
Heat Sources -..--.. ------- .......-.. ..... --- ... ..----.-.. 11
Performance of Indirect Expansion Reverse Cycle Refrigeration
Heating System .--. ....._ .. ... ............. .....--- ..--. 17
The Performance of Direct Expansion Reverse Cycle Refrigera-
tion Heating System - ......................... ..----. -- .... 22
Internal Combustine Engine Power for Reverse Cycle Refrig-
eration .__... _..... .. .... ..... ... - -- 24
The Economics of Reverse Cycle Heating .......-- ... .. -29
Novel Arrangements of Equipment for Reverse Cycle Refrig-
eration Heating ..-- ----...... ............... - .. 835
Heat Pump Installations ......... ........... -- -- 38
Conclusions ...----- --... . .............. .....- -.---- ....... 43
Glossary of Abbreviations ....-- .. ... .................. ....... 47
Bibliography -. -----.. .. .. ---..... ...------------- 48






REVERSE CYCLE REFRIGERATION FOR
HEATING IN THE SOUTH
Abstract: It is the purpose of this bulletin to show the present
state of development of the reverse cvcle of refrigeration system
for heating in the South-loth with electrical driven compressors
and with internal combustion engine driven compressors. Data are
presented which may facilitate the design and selection of equip-
ment for such applications. Since graphical methods of selection
of equipment are used, the operating costs under a definite set of
operating conditions can be determined. The prediction of seasonal
operating costs involve many variables and in general, it is recom-
mended that the operating costs be predicted by comparison of
the proposed installation \with an existing installation whose sea-
sonal operating cost is known. It is possible to make operating
cost computations on the basis of climatic data and heat losses
from the structure. Operating cost determinations based on this
procedure are shown in this bulletin. The examples illustrated are
for particular types of installations involving the use of definite
size equipment. However. the methods presented are applicable
to the determination of the performance of any size. type of equip-
ment. and installation which may be used.
Introduction: The reader might quizzically ask why consider
heating by the use of refrigeration equipment. Reverse cycle re-
frigeration heating cannot be consideel d economically feasible
unless there is a demand for summer air conditioning. If such a
demand exists. certain advantages may be obtained by the use
ot reverse cvcle heating. These are as follows:
1. The initial cost of equipment for year around air con-
ditioning is usually lower when reverse cycle heating is
used.
2. No flue or chimney is required with this type of heating
system when electric motive power is used. Therefore.
it is possible to save the construction cost of a flue or
chimney.
3. There is no smoke pollution problem when reverse
cycle heating is used, which results in the saving of
periodic interior and exterior decoration expenses, as
well as cleaner air in the neighborhood.
4. Often the use of this type of equipment permits a more
centrally located equipment room, thereby effecting a
saving in the heat distribution system.





5. For the electrically powered installation there is little fire
hazard when heating is accomplished by reverse cycle re-
frigeration. Therefore, this type of heating is particularly
adapted to certain types of applications.
6. Many times in a single building there is a need for sim-
ultaneous cooling of one or more areas in the building
with heating in other areas. When this condition exists
it is usually due to high internal heat gains in certain
areas which may be the result of occupancy, mechanical
equipment, electrical or gas appliances. Under this con-
dition reverse cycle refrigeration performs admirably as
heat is withdrawn from the areas to be cooled and de-
livered to the areas to be heated.
7. Frequently the cooling effect which is produced as a by-
product of the heating function can be used in an indus-
trial process which is carried on. within the building.
8. In some instances this type of electrical loading may be
desirable to the electric power companies and more fav-
orable rates may be granted to the owner of such a sys-
tem. This cannot be considered the general case as most
peak power demands occur during the winter months
or at the same time that the electric demand for reverse
cycle heating is at a maximum.
9. When electric powered reverse cycle heating is used
there is no necessity for fuel storage tanks or bunkers.
This storage of fuel often requires considerable useable
space within the structure.
The possibility of heating by the use of a refrigeration cycle
is not new. In 1852 Lord Kelvin *1 presented a paper showing the
possibility of heating with a refrigeration cycle; however, at this
early .date the equipment had not been developed to the point
where reverse cycle refrigeration could be used for heating. It was
in 1927 that the first successful reverse cycle heating system was
assembled-this application being made by T. G. N. Haldane2 in
the Southern California Edison Company's office building. Since
this application, considerable work has been done to perfect the
system; and at the present time there are many "tailor-made"
installations and there are some companies manufacturing factory-
built, self-contained, reverse cycle, heating units.
"Source of material as indicated In Bibliography.
2






A basic refrigeration system is shown in Figure 1. It is com-
posed of a compressor, condenser, liquid receiver, expansion valve,
and an evaporator with the necessary interconnecting piping. This
apparatus is connected in such a manner that during its operation
heat energy is transferred from one thermal potential to a higher
thermal potential and for this reason it is called a heat pump by
some writers. The refrigeration system as shown can be used as a
refrigerator or heater. If it is desired to use the system as a re-
frigerator, the heat absorbed by the evaporator is the useful
product. If the function of the system is that of a heater, heat is


BASIC MECHANICAL REFRIGERATION SYSTEM

QH= HEAT OUTPUT AT CONDENSER
QR= REFRIGERATING EFFECT AT EVAPORATOR
-= HEAT EQUIVALENT OF WORK INPUT TO COMPRESSOR

c< = REFRIGERATION ISOLATION OR SERVICE SHUTOFF
VALVES
FIG. I
3





taken away from the condenser and delivered to the area to be
heated. There are many ways of interconnecting the component
parts of a refrigeration system to achieve a heating effect. Primarily,
these methods fall into one of two categories:
1. Interchanging the function of the condenser and the
evaporator with no changes in the air distribution system.
2. Changing the air distribution system from the summer
cooling cycle to the winter heating cycle in order to de-
liver air from the evaporator to the outside and air from
the condenser to the inside space during the heating
cycle. In this type of application the refrigerant flow and
the function of each piece of refrigeration equipment is
fixed for both the heating and cooling cycle.
Due to the nature of the reversed cycle heating process, it is
evident that this type of heating should be considered only where
there is a demand for summer cooling as well as winter heating.
This is due to the fact that equipment and original installation
cost of refrigeration equipment to provide the heating function is
considerably higher than the cost of conventional heating systems.
But when summer air conditioning is to be provided, reverse cycle
heating is economical from the original cost point of view.

Thermodynamic Principles: The thermodynamic'principles of re-
frigeration and reversed cycle refrigeration for heating have been
presented by numerous authors. The following summary is in-
cluded in order to simplify the explanations used later.
In 1824, N. L. Sadi Carnot,: French engineer, published his
great work entitled, "Reflections on the Motive Power of Heat and
on the Machines to Develop that Power". In this publication,
Carnot proved that a cycle composed of a combination of iso-
thermal and adiabatic compression and expansion processes has
the highest thermal efficiency which can be obtained by any cycle
operating between the two given temperatures. Therefore, it is
used as a standard of performance to which actual cycles are
compared.
Figure 2 is a temperature-entropy diagram of this cycle. The
area ABCD of this diagram represents the work of compression of
a Carnot refrigeration system. Area ABC'D' represents the heat
dissipation at the condenser of the refrigeration system, assuming
that there is no loss by radiation, etc. Area CDD'C' represents the
cooling or refrigeration effect which is obtained. It can be shown
that the refrigeration effect and heat output at the condenser ex-






ceeds the heat equivalent of the work input to the compressor
when the difference between Te and Te is less than the distance of
Te from absolute zero. It is for this reason that reverse cycle re-
frigeration for heating has commanded such an interest from engi-
neers. The ratio of the refrigeration effect to the compressor input
is termed the coefficient of performance of the system as a
refrigerator.


-


Dr C' s
TEMPERATURE ENTROPY DIAGRAM
OF A CARNOT REFRIGERATION CYCLE

AREA-ABCD- WORK INPUT TO COMPRESSOR
AREA- CDD'C'- REFRIGERATION EFFECT
AREA- ABC'D'- HEAT OUTPUT AT CONDENSER
FIG. 2
5


-%1-------1





T. -Equation 1.
C T-T.
Where CP, is the coefficient of performance as a refrigerator.
T. is the refrigerant evaporating temperature expressed in
degrees absolute (OF+460").
Tc for the refrigerant condensing temperature in degrees
absolute (*F+4600).
The coefficient of performance of the reverse cycle of the
refrigeration system (Carnot cycle) when used for heating is CP h.
T. -Equation 2.
CP h T
I -T.
Since the two equations above are based on Carnot's cycle, or
ideal performance, it is impossible in actual practice to obtain
coefficients of performance as high as those indicated by the use
of these two equations. In practice it is impractical to duplicate
the reversible isothermal and adiabatic processes which were as-
sumed by Carnot.
A practical refrigeration cycle can be shown on a pressure-
enthalpy diagram and the changes of pressure and enthalpy for
the various processes may be clearly illustrated. Figure 3 is a PH
diagram for a Freon-12 refrigeration system operating at a con-
densing temperature of 1020 F and a refrigerant evaporating
temperature of 30 F. The refrigerant leaving the evaporator coil
is superheated 100 F; the refrigerant entering the expansion valve
is subcooled 100 F. On this diagram the process indicated by line
AB is the adiabatic throttling of the liquid Freon through the ex-
pansion valve. The process BC is the evaporation of the wet vapor
from state B to the superheated state C. Process CD accounts for
the heat energy input to the system by the compressor motor. This
heat addition is due to the fact that the compressor motor is not
10095 efficient. It should be noted that process CD is present on
the PH diagram only when a hermetically sealed refrigeration
compressor is powered by an electric motor and when the motor
is cooled by the suction gas from the evaporator.4 Process DE is
the compression process from the superheated state D to the high
pressure superheated state at point E. In an ideal compressor, the
compression process would be adiabatic and would follow a line
of constant entropy. However, due to the friction, throttling, and
other mechanical causes, it is impossible to compress the refrig-
erant adiabatically. Line DE represents a polytropic compression
process which is followed by the actual compressor. This polytropic












idO tSTAN TE MPEIRATU

oo


-P 3




1, 0C


- --
I















compression is represented algebraically by the following equation.
Pvn =Constant -Equation 3.
Where n =exponent for polytropic compression
P=absolute pressure in lbs. per sq. ft.
v=specific volume in cubic ft. per lb.

Process EA is followed in condensing the refrigerant in the
condenser. In this process, heat is released from the refrigerant
through another medium which may be air in the case of air
cooled condensers-or water in the case of a water cooled
condenser.
It may be noted from Figure 3 that heat energy is added to
7
IX J S_____ ___ __ -
- - - T ^ ^ ^ r
I "^s. (
_ ____ ___ ____ __ ___ L -- ----- ^,- --- --- -*





the refrigerant from two sources: (1) the compressor motor, and
(2) the refrigerant evaporator. Heat is released from the system
at the condenser. It is obvious that there must be a source and
a receiver of heat energy in order for the coefficient of performance
of the refrigeration system to be greater than one.
Table 1 is a compilation of data on an actual refrigeration unit
showing the refrigeration effect, heat dissipation at the condenser,
and electrical input to the compressor.4 The thermodynamic co-
efficients of performance which are obtained in practice are also
shown. This table covers the range of several refrigerant evap-
orating temperatures and several condensing temperatures.
Most authors on the subject of reverse cycle refrigeration for
heating have confined their discussions to the use of electric motive
power for the compressor. It can be shown that lower operating
costs can be obtained by the use of internal combustion engines
for motive power. The use of such equipment for motive power
can in some instances nullify a portion of the advantages obtained
by the use of reverse cycle heating systems. The motive power
for this type of heating system should be carefully considered
before definite conclusions are reached. In the application of in-
ternal combustion engine driven heat pumps, practically all of the
heat energy of the fuel may be utilized. At the same time the
coefficient of performance of the refrigeration system increases the
total heating capacity of the system to a value in excess of the
heating value of the fuel consumed. This particular type of appli-
cation is discussed in detail under the section entitled, "Internal
Combustion Engine Motive Power".
From the data given in Table 1, it is possible to plot Figure
4 from which several conclusions can be drawn.
1. The output from the refrigeration system increases as
the refrigerant evaporating temperatures increase for a
constant condensing temperature.
2. For a fixed refrigerant evaporating temperature the heat
dissipated to the condenser increases as the condensing
temperature decreases.
3. The electrical energy consumed by a given compressor
operating at a fixed speed increases as the refrigerant
evaporating temperature increases. However, the ratio of
heating effect to the compressor input (CPh) increases
as the refrigerant evaporating temperature increases.
Also, it can be seen that CPr increases with an increase
in refrigerant evaporating temperature.






TABLE No. 1
COMPRESSORFERFORMANCE OF A FREON 12 COMPRESSOR AT VARYING REFRIGERANT EVAPORATING
TEMPERATURES AND CONDENSING PRESSURES-870 RPM-20 HP COMPRESSOR
100 PSI-90'F. CONDENSING TEMPERATURE

Motor Input
Refrigerant Evaporating Motr Heat from Evaporator Actual Heat from Condenser Actual Carnot Cycle
Temperature *F. KW BTU/HR BTU/HR CP, BTU/HR CPh CPb

10. .. .. 12.9 44.040 104.000 2.35 148.040 3.35 7.01
14. 13.3 45.400 116,000 2.55 161.400 3.55 7.37
18. 13.7 46,800 130,000 2.79 176.803 3.79 7.78
22. 14.0 47,800 145,000 3.03 192.800 4.03 8.24
26... 14.2 48,500 162.000 3.35 210.503 4.35 8.75
30- -. 14.4 49,200 179.000 3.65 228,200 4.65 9.33
34. 14.5 49.500 197,000 4.02 246,500 5.02 10.00
38. 14.5 49,500 218,000 4.50 267.500 5.40 10.78
42. 14.5 49,500 241,000 4.86 290,500 5.86 11.68
46. .14.4 49,200 265,000 5.40 314,200 6.40 12.72

130 PSI- 107"F. CONDENSING TEMPERATURE

10...... 13.1 44.700 38.000 1.97 132,700 2.97 5.85
14... 13.7 46.800 100,000 2.14 146.800 3.13 6.10
18. 14.3 48,800 113,000 2.32 161,800 3.31 6.37
22. 14.8 50,600 126,000 2.49 176,600 3.49 6.67
26. 15.3 52.300 141,000 2.70 193.300 3.70 7.00
30. 15.7 53.600 157.000 2.93 210,600 3.94 7.37
34 16.0 54.700 174,000 3.19 228,700 4.18 7.77
38 ......... ....... ... 16.3 55,700 192,000 3.45 247,700 4.44 8.22
42. ... 16.5 56,400 213,000 3.78 269,400 4.77 8.73
46. ................. 16.6 56,700 234,000 4.13 290.700 5.14 9.30
50. .... ....... 16.7 57,100 256,000 4.48 313,100 5.48 9.95














U
h 3130 PSI


00

300000
S 100 PSI
00 PSI / CONDENSING
PRESSURE
nEAT DISSIPATED
FROM CONDENSER /13PSI





200000





HEAT ABSORBED
-- BY EVAPORATOR






N130 SI




.5 NO
aht 00 Posbls a
f10 g



0 0
10 14 18 22 26 30 34 38 42 46 50
REFRIGERANT EVAPORATING TEMPERATURE
20 HP CONDENSING UNIT F-12 870 RPM
rIG NO. 4


Therefore, for maximum efficiency the compressor should oper-
ate at as high a refrigerant evaporating temperature as possible and
at as low a refrigerant condensing temperature as possible.






Heat Sources: In the normal refrigeration cycle the source of heat
is the useful cooling of the desired area or product: and the receiver
of the heat is the condenser cooling water in the case of a water
cooled condenser, or air if an air cooled condenser is used. When
reversed cycle refrigeration is used for heating, the source of heat
is the medium which is cooled by the evaporator, usually water
or air. and the receiver of the heat is the building to he heated.
It was shown in Figure 4 that for maximum efficiency either as a
refrigerator or as a heater the refrigeration system should operate
at a high refrigerant evaporating temperature and a low refrigerant
condensing temperature. The actual \alues for these temperatures
are set by other engineering considerations such as the tempera-
ture of the heat source and the temperature to be maintained in
the heated space. Also. to be considered is the economic problem
of the cost of heat transfer surfaces.
The two basic heat sources for reverse cycle refrigeration heat
are water and air. However. latest information indicates that the
earth may be used as a heat source." When water is available and
economical to use it will he found that reversed cycle refrigeration
for heating is not only feasible but very practical. There are several
considerations which must be taken into account when the use of
water as a heat source is contemplated. These considerations are:

TABLE No. 2
MAXIMUM. MINIMI'M TEMPERATURES, WATER COSTSS AND DISPOSAL
REGULATIONS FOR SOUTHERN MIL NIIPALITIES

Ci r Min Tcmp. Max. Temp C(ost Regulation
Birmingi.ii A. 48 52 i 5.(X None Gi\cn
Mobile. Ala. 40 80 23.25 None2
Gaincsvillc. Fla.. I 7S 84 13.08 Sec Cir Engr.
Jacksonille F!a. 78 84 9.25 None
Miami. Fla.. 75 78 14.93 YesI
Orlando. Fla 60 90 18.00 Yes
Tampa. Fla 70 I 86 19.90 See (itv Engr.
Atlanta. Ga 33 87 33.80 IYes 3
Macon. Ga. 40 70 Sec Cirv Engr
Savannah. Ga. 72 72 14.84 None Given
Newv Orlc.ua. La 42 91 10.60 Yes
Sparanhurg. C 40 78 26.97 None
Dalla,. Tex 39 86 24.00 None"
San Antonio. Tex. 76 81 14.04 Yes:


1. Cost Figured for First 100.000 Gallons per Month
2. Regulation Contemplated in Near Future
3. Regulation Contact City Engineer
4. Discharge Condenser Water to Storm Sewer
5. Depends on Installation. Stricter Regulations Contemplated
6. Discharge Condenser Water in Storm Sewers if Possible.






(1) Adequate supply
(2) Adequate water disposal system
(3) Cost of water
(4) Temperature of water.
Since the water for the heat source is to be obtained from
municipal utilities, wells, lakes and rivers, it becomes a simple
matter to estimate the cost involved. In most instances where water
is to be used it will be found necessary to purchase it from munic-
ipal utilities. Table 2 shows the cost for water from various
municipal plants in several of the Southern cities. Also included
in this table are data on the observed minimum and maximum
temperatures of the water. In estimating operating costs it should
be noted that usually the average winter water temperature is
considerably higher than the minimum water temperatures given
in this table.
Table 3 gives the monthly temperature readings from two
Southern municipal water supplies.
TABLE No. 3
AVERAGE MONTHLY WATER TEMPERATURES
FROM MUNICIPAL WATER SUPPLIES

Spartanburg, S. C. New Or'cans. La.

Month Muyimum Minimum M x mum Minimum

January 55 40 72 42
February 48 40 69 51
March 61 48 69 51
April 65 58 77 59
May 69 60 82 70
June 76 68 89 77
July 78 74 95 83
August 76 72 91 84
September 91 78
October 89 72
November .. 78 57
December . 69 50

Figure 5 is a map showing the water temperatures from non-
thermal wells throughout the United States."
The quantity of water required depends upon the allowable
temperature drop of the water passing through the evaporator, the
coefficient of heat transfer and the total heat to be transferred from
the refrigerant to the water. The quantity of water consumed is
based on engineering economic considerations which involve the
solution of two equations.








o -










































.L





= Refrigeration elect huri per hour
=Area of the evaporator surface sq. ft.
-The cocfticienr of hear transfer Bru per sq. fCr. hr. F
= Logarithmic mean celcctive temperature diffcrcnce-F


\Vhere Q,
A
U
t,,,








gpm --Equation 5.
(t.-tl) x 8.3 x 60
Where gpm=gallons of water passing through the evaporator per
minute
tI =temperature of water leaving evaporator F
t: = temperature of water entering evaporator OF
Q, = refrigeration cffect-Btu per hour
Engineering performance data are available on the performance
of evaporators and condensers. By referring to these data from
the manufacturers of the equipment it is possible to determine the
proper amount of water to be circulated without the use of equa-
tions 4 and 5.
When air is used as the heat source, it will be found that the
system usually will operate at a lower efficiency than when water
is used. This is caused by the low temperature of the heat source.
l)Data included in Table 4 shows the design outside air temperatures
TABLE No 4
CLIMATIC DATA FOR DESIGN OF
AID CONDITIONING SYSTEMS IN FLORIDA '

Lowest Highcst
Temperature Wintcr Deign Tcrmprature Summer Dc;:gn
(Cv n on Record d F J on Record F db F ah
F db 'F wb

Jackson~ e .1 10 25 20 104 96 78
Miami. 27 30 26 96 92 80
Pcns.cola 7 20 17 103 98 81
Tamp; 19 30 26 98 96 80

Design temperatures not exceeded over two per cent of days during 10 year period
1927-1937.

for heating and cooling in Florida. For other locations the reader
is referred to the American Society of Heating and Ventilating
Engineers' Guide. The generally accepted practice in the determin-
ation of design temperatures is not to select the most extreme
temperature ever recorded for the locality. This selection would
indicate the installation of equipment which might operate at peak
capacity only once in a 25 to 50 year period. Therefore, it is general
practice to investigate the climatic data for a ten year period and
to select temperatures which are not exceeded during any hour
of the day on more than two per cent of the days. This means that
98 per cent of the days the system is operating at or less than full
capacity but since the extreme temperatures during the winter
14






occur during the early morning hours when the heating system
does not have to fulfill exact inside design conditions, the heating
system sized by sichl design temperatures is adequate \even on the
remaining two per cent of the days. Also. it should be noted that
building structions have a thermal capacity and a heat transmission
lag which tend to minimize tile helects of excessively low temper-
atures of short duration.
It should be noted that the data given in Table 4 cannot lie
used for estimating operating costs as this table is for determining
the capacity of systems to be installedd. '. Table 5 is a typical
table to be used for determining operating cost. Average normal


SUPPLY AIR

XNSO RETURN AIR
EXPANSION TANK |


OPEN SUMMER CYCLE
OPEN WINTER CYCLE


INDIRECT EXPANSION AIR CONDITIONING SYSTEM ARRANGED FOR
REVERSE CYCLE HEATING

FIGURE 6






TABLE No. 5
CLIMATIC DATA FOR AIR CONDITIONING -JACKSONVILLE


A-AVlRACG NORMAL DAYS

Ilii. J., I FI. IMA. Aa. M. MAY JUNI JULY ( Auo. SEP. Ocr. |Nov. I Dc.

Dayv BULB TiuPERATURE IN *F

I s5.J 56.0 54.6 63.4 70.1 73.7 774 76.8 75.6 70.3 6o.6 56.5
4.t6 55.7 54.5 63.6 69.7 73.2 76.8 76.8 75.5 7o.o 6o.o 55.3
1 54.- 54.9 53.8 63.1 69.1 73.1 76.2 76.0 74.9 69.6 59.2 54.3
4 54.1 54.9 54.1 62.3 68.6 73.2 75.8 75.8 74-7 69.1 59.4 53.7
5 53.6 54.7 53.9 61.6 68.0 .3.0 75-5 75.7 74.6 69.3 59.1 51.8
6 53.4 54.2 54.1 61.6 67.3 73.3 754 755 74-.3 68.7 58.9 52.7
7 53.2 54.0 54.1 61.6 68.9 75.4 76.6 76.6 74.8 68.6 58.o 5.0
5 53.5 54.2 54.9 64.5 71.5 76.8 78.8 79.3 79.2 69.4 58.3 521.
a 53.2 SS.8 56.0 67.8 75.0 79.8 81.1 82.4 8:.6 69-7 59.4 53.9
Io 54.3 58.t 59.3 70.2 77.4 82.1 82.6 84.2 82.1 72.3 62.0 56.4
I 56.4 61.7 61.4 71.- 79.7 84.1 84.9 85.8 83.1 73.6 63.7 58.3
Nin 57.3 63.5 63.8 73.4 81.1 85.2 85.9 85.9 j3.o 74.7 64.5 59.6
I 59.1 65.7 65.0 75.1 8S.0 86.8 87.0 86.1 83.2 74.9 65.7 6o.3
59.9 65.6 65.8 75.7 82.5 87.2 87.3 86.5 82.7 74.6 65.9 60.7
60.8 65.4 66.1 75.6 82.1 86.3 86.0 85.4 82.7 74.5 66.3 6s.o
I 60.3 65.1 66.0 75.1 80.5 83.8 83.5 84.9 82.2 73.5 66.1 60.8
I 59.2 64.0 65.1 74.6 78.3 81.S 82.6 83.5 81.2 73.0 65.3 60.3
1i 56.9 61.8 62.7 73.0 77.3 77.5 82.o 82.4 80o. 7z.2 63.8 58.8
; 54.8 59.5 61.4 70.9 75.1 76.6 81.t 81.2 78.6 71.1 624 574
53.7 57.8 60.5 69.1 74.7 76.3 79.1 80.3 77.6 70.3 61.9 56.8
9 52.2 56.8 59.8 67.5 73.3 75.6 78.5 79.2 77.2 69.6 6l.4 55.6
Io 51.5 56.5 S9.5 67.1 72.1 75.3 78.1 78.3 76.7 69.7 61.3 55.4
11 51.0 56.1 59.0 66.9 71.4 75.S 77.5 78.1 76.5 69.S 6o.7 54-4
12 50.6 SS.7 58.5 66.8 71.0 75.2 77.4 78.1 76.0 68.7 60.0 54.6
Av. 55.1 58.6 59.3 71.0 74.4 78.3 80.2 80.6 78.6 71.1 61.8 56.4
WET BULB TEMPEIATUR IN *F

8 n.m. 49.0 48.5 51.6 59.6 67.4 72.3 74.- 74.9 72.5 65.5 54.8 48.0
2 N. 48.2 53.4 56.0 64.0 68.9 74.3 75.1 75.0 73.7 65.6 58.8 st.6
sp.nMl.45.o 52.8 54.6 63.8 68.3 71.8 74.3 74.6 73.o 65.3 58.6 51.6


B-AVEMAGS ExTaimu Dayg

HI. JAN. FS. JULY Auc. Sa. Dac.

Day BULB Tv.MPEIIATUR IN F

1 40.0 41.0 79.1 79.6 76.7 41.7
2 39.0 39.5 78.9 79.o 77.3 40.5
S 38.0 38.J 78.4 784 76.9 39.1
4 J6.6 36.0 78.1 77.8 76.3 38.4
5 35.7 .5.6 77.9 77.5 75.8 37.8
6 34.7 33.8 77.6 77.3 7S.S 384
7 34.1 3J.3 78.2 77.9 76.1 J38.
8 34.1 33. 80. o., 77.5 j8.t
9 14.6 35.3 85.o 8j.3 83.3 39.5
so 38.0 j8.8 86.0 86.1 84.0 41.6
11 41.7 42.4 87.8 88.7 87.1 43.6
Noon 44.8 45.2 89. 90.5 88.7 44.-
3 46.9 47.9 89.7 91.5 89.5 47.5
2 48.5 48.6 90.4 92.7 89.6 484
S 48.J 50.0 90.8 93.1 90.6 49.5
4 48.4 50.0 89. 9314 88.1 49.3
5 47.9 49.8 88.9 89.: 86.3 47.9
6 46.0 48.0 874 86.6 86.3 46.3
7 44.8 46.6 86.2 85.o 8.o0 45.6
8 43.9 46.0 84.S 82.6 80.o 454
9 4J.O 46.2 82.S 81.4 79.0 44.8
o0 42.3 46.2 81.8 o0.9 78.5 44.1
I 41.6 46.2 80.9 80.5 78.1 43.4
33 40.5 46.0 80.6 79.8 77.7 43.9

WET BULB TIMPRATUml IN *F

8a.m. 31.o0 3.5 75.0 75.o 73.t J5.S
12 N. 37.5 15.0 76.0 75.3 75.o J6.2
Sp.m. 38.0 41.5 75.2 75.0 74.3 39.0





temperatures which are experienced by four cities in Florida may
be obtained by reference to Bulletin No: 5 of the Florida Engineer-
ing and Industrial Experiment Station entitled, "Climatic Data for
the Design and Operation of Air Conditioning Systems in Florida".
Experimental work to determine the possibilities of the earth as a
heat source and heat receiver is contemplated by this Station and
the results of these tests will be published at a later date.

Performance of Indirect Expansion Reverse Cycle Refrigeration
Heating System: The diagrammatic sketch, Figure 6, shows a
method of connection for an indirect expansion reverse cycle re-
frigeration heating system including the summer winter change
over valves required. The change over valves may be either manual
or automatic. The graphical solution to determine the performance
of indirect expansion refrigeration systems for reverse cycle re-
frigeration heating is one of the simplest methods of solving prob-
lems involving operating balance and selection of proper equip-
ment. For this reason the method is recommended.
Figure 7 is a graph showing the performance of a 20 hp, her-
metically sealed, six cylinder, Freon-12 refrigeration compressor
driven by an electric motor at 870 rpm with a matched water
cooled condenser and a matched indirect expansion evaporator.
Data to plot this graph are taken from manufacturer's performance
data on the equipment. This graph shows the effect of varying the
entering water temperature to the condenser, the refrigerant evap-
orating temperature, the heating output from the condenser, the
refrigeration effect of the evaporator, the temperature of the water
leaving the evaporator and the temperature rise and fall through
the evaporator or the condenser for a given capacity and water
flow.4 By the use of the graph it is possible to determine the per-
formance of this matched refrigeration system under various con-
ditions. The results of such solutions are shown in Figures 8, 9,
10, 11 and 12.
Figure 8 shows the heating effect obtained from this indirect
expansion system for various temperatures of water leaving the
condenser. The inlet water temperature to the evaporator is held
constant at 65' F and the volumes of water flow through the
condenser and the evaporator are identical. The heating effects
shown on this figure are for two rates of water flow through the
condenser and evaporator. These rates of flow are 1800 gph and
2400 gph. Interpolation may be used for intermediate water flows.
Figure 9 indicates the heating effect obtained at varying inlet


















t wt t


w e t t ev orto s held t h Ig
F Io'mhese ui i.r








1i

.1.
| 5\ I, 11

I ti r i r t o gi













water temperatures to the evaporator for refrigerant condensing
temperatures of 96' F, 1070 F and 1170 F when 2400 gallons of
water per hour are passed through the evaporator.
Figure 10 shows the heating capacity of the same system with
varying inlet water temperature to the evaporator and refrigerant
condensing temperatures of 960 F, 1070 F and 1170 when the
water to the evaporator is held at the rate of 1800 gallons per
hour.
From these curves it can be seen that as the inlet water tem-
18
e~~~~~~ g^^\^ ^_ 3 __ ; \
r~^^^t Si I H g
p ^ ~ ~ '" g ~:p t
















18





400 000


3301

0



200

I-
I-
j


70 80 90 100 110
TEMPERATUI OF WATER LEAVING CONDENSER 't

HEATING EFFECT VS TEMPERATURE OF WATER TO HEATING
COIL -INDIRECT EXPANSION-WATER TO EVAPORATOR 65-F
FIGURE 8

perature to the evaporator is decreased, the capacity of the system
as a heater is decreased. From Figure 8 it can be seen that the
capacity of the system increases with a constant inlet water tem-
perature if the condensing pressure decreases. There is a practical
limit to which the condensing temperature can be decreased. This
being a function of the size of the heat transfer surfaces used to
warm the air delivered to the space to be heated. In general, it will
be impractical to use condensing temperature lower than 960 F.
This is due to the temperature differentials required to warm the
air sufficiently for heating.
Figures 7. 8. 9 and 10 have shown the performance of the
refrigeration equipment as previously listed under various con-
ditions; however, it is necessary to also match the performance of
the water coil in the air conditioning unit with the refrigeration
system's heating capacity in order to determine the actual amount
of heat which may he realized for heating purposes. Included in


2400 cOHN T7rOUG CONHOMSER





IBOo GC TOt,&CH LVAPORATOI


000 _





000





0
ooo _____________






FIGURE 9


300000






200 000

0
I


D:
a-

- 100 000



0


0


9 107
96 I07


1
U


40 50 60 70
TEMPERATURE WATER ENTERING
EVAPORATOR -*F
HEATING EFFECT AT VARIOUS CONDENSING
TEMPERATURES FOR WATER HEAT SOURCE

Figure 11 is a curve which shows the heating effect available from
the condenser of the matched refrigeration equipment previously
listed when the rate of water flow through the condenser is 1800
gph and the water flow through the evaporator is 1800 gph. The
heating effect is plotted against water temperature leaving the
condenser. This curve is plotted from data given on Figure 8. Since
the rate of water flow through the condenser is known and the
temperature of the water leaving the condenser and entering the
water coil is known, the capacity of a given water coil for various
entering air temperatures may be computed.10 For this particular
application the size of the water coil and total air circulated were
established by summer air conditioning requirements. These were
as follows:
Water coil 15 sq ft face area 5 rows deep
Total air 7500 cfm
Fresh air 1500 cfm


117


EVAP. 2400 GPH
COND.2400 GPH






300000





D 200000
0

aJ

SI
(a 100000


a:
96 1071

517













EVAP. 1800 GW-
COND 1800 GPM



10 50 60 70
TEMPERATURE WATER ENTERING
EVAPORATOR-*F
HEATING EFFECT AT VARIOUS CONDENSING
TEMPERATURES FOR WATER HEAT SOURCE
FIGURE 10


Inside temperature-winter 70 F
Outside design temperature-winter 20' F.
Under these conditions the limiting temperatures at which air
can enter the water coil is 60' F and 70' F. Therefore, if the per-
formance of the water coil is plotted on Figure 12 as shown for
a 600 F and 70 F entering air conditions, the limits of the opera-
tion of the system are set. For intermediate temperatures of air
entering the evaporator, the heating capacity can be determined
by interpolation.












HEAT OUTPUT
rno cowo -
WATER COIL
s0 AS A HEATER






30-





9O 100 Io 0 10 Ia 0
TEMPERATURE OF WATER LEAVING CONDENSER AND ENTERING TEL
I-EATING COIL-Fr.

PERFORMANCE INDIRECT EXPANSION SYSTEM HEATING CYCLE

FIGURE II
From Figure 11 data is supplied to plot Figure 12 which shows
the heating capacity versus outside air temperature for the reverse
cycle of refrigeration system. This refrigeration equipment is com-
posed of a 20 HP, 870 rpm, six cylinder, hermetically sealed, electric
motor driven "Freon 12" compressor with its matched indirect ex-
pansion water chiller, and a water cooled condenser. The flow
through condenser and water chiller is 1800 gph each. A water
coil of 15 sq ft face area and 5 rows deep is installed in the air
conditioning unit. The air conditioning unit has a total air flow of
7500 cfm of which 1500 cfm is fresh air. Also shown on Figure 12
is the electrical input to the compressor and the water pump. The
fan power is not included as this power would be required if a
conventional warm air furnace were used.

The Performance of Direct Expansion Reverse Cycle Refrigeration
Heating System: Figure 13 shows a schematic diagram for inter-







Scors Epur. \_ _( --
COMP OMLTY







0


s0 40
OUT3j10 AIR TEuPCRATURE *r


OVERALL PERFORMANCE OF INDIRECT EXPANSION SYSTEM ON THE
HEATING CYCLE AND ELECTRIC POWER CONSUMPTION VS OUTSIDE
AIR TEMPERATURE
FIGURE 12
connection of a direct expansion type system with the necessary
switch-over arrangement from summer cooling to winter heating.
The performance of the direct expansion reverse cycle heating
system can also be plotted graphically and the performance at
various specified conditions determined from this graph.10 Figure
14 is a graphical representation showing the performance of a
direct expansion reverse cycle refrigeration heating system.
Figure 15 shows the overall performance as a heater of the
direct expansion system for varying entering air temperatures to
the evaporator.
For the particular example shown, it will be noted that the
condenser selected is slightly larger than would normally be sup-
plied for summer cooling. Oversize selection is not necessary; how-





TO OUTSIOC


AIr


,Ma K I| L.U...l -. I-....-.. i-i i
C.*LLE CYLLE
sUPO#4 W4AN or -AD UPON LIAW(.C or


ULtKIW FAN




ALL DAMPER POSITIONS SHOWN
FOR wiNtER CYCLE OPERAriOn
EFRILMRANT FLOW JIDNT:LAL
FOR IOTWH CCLCS

COMPRESSOR


DIRECT EXPANSION AIR CONDITIONING SYSTEM ARRANGED FOR
REVERSE CYrLE HEATING
FIGURE 13

ever, such procedure results in greater heating capacities and lower
condensing temperatures which are used in the interest of obtaining
reduced operating costs.

Internal Combustion Engine Power for Reverse Cycle Refrigeration:
The discussion of reverse cycle refrigeration for heating has been
confined thus far to installations using electric motors for prime
movers. A gain in thermal efficiency by the use of internal com-
bustion engines as prime movers is possible. From an unbiased
viewpoint it must be acknowledged that most of the inherent ad-
vantages of the electric powered reverse cycle system as enumer-
ated in the introduction are sacrificed to obtain this gain in thermal
efficiency.
The Internal combustion engine driven heat pump has basically
the same heating capacity as illustrated in the foregoing sections.
The overall heating capacity of the refrigeration cycle is reduced






CONO TEMP.
96* IOlrIit


10 20 30 40 SO
REFRIGERANT EVAWPOATING TEMPERArURE *F

DIRECT EXPANSION AIR CONDITIONING SYSTEM PERFORMANCE
AS A REVERSE CYCLE HEATING SYSTEM- 20 HP. COMPRESSOR
FIGURE 14

approximately 0.5 per cent since the foregoing sections assumed
the use of a hermetically sealed refrigerant cooled electric com-
pressor motor. When using this type of refrigeration compressor,
nearly 100 per cent of the electric energy delivered to the com-
pressor motor plus the refrigeration effect is available for heat at
the condenser of the refrigeration system. On conventional open










COMP. OMLY

I-






-"E
% .

























uixID AIR
HEATING CAPACITY VARIATION WITH AIR TEMPERATURE
FOR 20 HP COMPRESSOR DIRECT EXPANSION 7500 CFM
THRU EVAPORATOR AND 7500 CFM THRU CON-NSER









FIGURE 15

type refrigeration compressors, which would be used for internal
I-























combustion engine driven heat pumps, not all of the energy of the
prime mover is available at the condenser. This accounts for the
loss of about 0.5 per cent of the heating effect.

A typical Diesel engine's performance is shown on Figure 16.
These curves indicate the thermal efficiency of the Diesel engine
at varying loads, the energy consumed to overcome friction heat
dissipated to the cooling water, and the heat released to the exhaust
gas at full and fractional loads. By using internal combustion engine
z
I-







40 50 6O 70 60
TEMPERATURES OF AIR ENTERING EVAPORATOR -*l-OU1SIDE OR
MIXED AIR
HEATING CAPACITY VARIATION WITH AIR TEMPERATURE
FOR 20 HP COMPRESSOR DIRECT EXPANSION 7500 CFM
THRU EVAPORATOR AND 7500 CFM THRU CONDENSER
FIGURE 15

type refrigeration compressors, which would be used for internal
combustion engine driven heat pumps, not all of the energy of the
prime mover is available at the condenser. This accounts for the
loss of about 0.5 per cent of the heating effect.
A typical Diesel engine's performance is shown on Figure 16.
These curves indicate the thermal efficiency of the Diesel engine
at varying loads, the energy consumed to overcome friction heat
dissipated to the cooling water, and the heat released to the exhaust
gas at full and fractional loads. By using internal combustion engine






100 Radi *on te Lo e




80















Net Effective Wor



20-
/--

/==
0. - - - _


50
Percent of Load


Full
Load


HEAT BALANCE OF A DIESEL ENGINE
FIGURE 16

power for reverse cycle heating, the heat energy of the cooling
water and exhaust gases may be utilized.
A proposed internal combustion engine powered reverse cycle
refrigeration system is shown in Figure 17. It will be noted that
the cylinder jacket cooling water is pumped from the engine
through the water coil in the air conditioning unit placed behind
the condenser; heat energy from this source is usually available


o


P1
E-i


I4






CwHAUST GAS
MEAT EXCA.UGCC


I


TO OUTSIDE


TO OUTSDOC


TM$( DAMPERS OTATE LIN TIrsr DAMPERS *OTATE
so* UPON ANGC Or CYCLE i )* UPON CNANG or
TO oUTSIDE oc

REC VER PAN

SUCTION GAS
LINE LINE


ALL ODMPERS SMON IN WINTER
CYCLE IjO ON.
VALV OPEN WINTER CYCLE O.LY
LV OPEN aSUNI CCLE l 0NLY
RfRIlGERANT fLOW IOENTICCAL O0
-iOTH CYCLES
C4.r.ET O SE11S COMPaRSSOR
WATER ENGINE
PUMP t I W


DIRECT EXPANSION SYSTEM POwERED WITH AN INTERNAL COMBUSTION
ENGINE ARRANGED FOR REVERSE CYCLE HEATING
FIGURE 17

at about 160 to 180" F. This system permits the condenser of the
refrigeration system to have unimpaired capacity while an eco-
nomically sized water coil for cooling the jacket water may be used.
It will also be noted that the refrigeration system proposed is of
the direct expansion type which usually has a lower initial cost of
installation. The system is equally applicable to the use of indirect
expansion refrigeration systems. The percentage gain in capacity
is essentially the same as that indicated by the analysis for the
direct expansion system. The direct expansion refrigeration system
uses the exhaust gases to preheat the air entering the evaporator
during the winter cycle. This enables the refrigeration system to
operate at more satisfactory refrigerant evaporating temperatures.
The exhaust gases should pass through a heat exchanger and not
directly through the evaporator as oil and carbon residue will col-
lect on the evaporator surface and impair the coefficient of heat
transfer. The resistance to flow of exhaust gases should be held to





a minimum as increased back pressure on the internal combustion
engine reduces its efficiency. The use of an exhaust gas heat ex-
changer should always be confined to the air circuit which is de-
livered to the outside as in this way the danger of delivering carbon
monoxide to the conditioned area is avoided. The exhaust gas heat
exchanger is provided with valves to direct the flow through a
conventional muffler during the summer cycle of operation.
The heating effect which can be obtained from an internal
combustion engine powered reverse cycle direct expansion system
is shown on Figure 18 for varying outside air temperatures. The
refrigeration system used in computing data for this curve is sim-
ilar in all respects to that used in computations for Figure 15 except
that an open type refrigeration compressor without a refrigerant
cooled motor is used. The second curve on Figure 18 shows the
specific fuel consumption per hour when operating at varying out-
side air temperatures.

The Economics of Reverse Cycle Heating: From the curves pre-
viously presented, it is possible to compare the cost of operation
of the following types of heat pumps; electrically powered indirect
expansion systems, electrically powered direct expansion systems
and internal combustion engine powered direct expansion systems.
The operating costs per therm (100,000 Btu), not including invest-
ment, interest depreciation, service or maintenance, for the three
systems previously discussed are shown in Figure 19 for varying
electrical energy and fuel oil rates. Also shown on Figure 19 is the
operating cost of an oil fired furnace burning 18,000 Btu lb oil at
75 per cent efficiency at varying fuel oil costs.
This, however, is not the true operating cost of these systems
as the original installation investment, depreciation, interest, and
service and maintenance must be considered in arriving at a true
conclusion as to overall operating and owning costs. This question
can be answered by the solution of Equation 6, by which the actual
cost of owning may be compared.
To compare the cost of an air conditioning system on a fair
basis, it is necessary to compute the total cost involved which in-
clude initial investment, interest on the investment, depreciation
of equipment, electrical energy required to operate the system
and or the fuel required to operate the system as well as the cost
of servicing the equipment. Equation 6 takes these factors into
account:








2


9







01
8
.J




400 000



























0
4 3ooo-o





















40 so 60 70
IE PERATURE Of OUTSIDE OR MIxtD AIR ENTERING
EXHAUST GAS HEAT EXCHANGER 'r

HEATING CAPACITY AND FUEL CONSUMPTION FOR
DIRECT EXPANSION HEAT PUMP WITH DIESEL ENGINE
MOTIVE POWER
K


























FIGURE is

IC
++OC+SCActual Yearly Owning0Cost -Equaion 6.














Where IC is the initial cost of a summer air conditioning system
and the heating system.

30
10000 -






0 -----------------------------------
40 SO 60 70 SO
iCUPIRATURE Or 0UTSIDC OR MIX*W AIR ENTCRINC
CXIHAUST OAS HIAT ?.CCHANCER -r

HEATING CAPACITY AND FUEL CONSUMPTION FOR
DIRECT EXPANSION HEAT PUMP WITH DIESEL ENGINE
MOTIVE POWER
FIGURE 16


-+I+OCSC= Actual Yearly Owning Cost -Equation 6.

Where IC is the initial cost of a summer air conditioning system
and the heating system.

80





COMPARATIVE COST OF FUEL AND ELECTRICITY FOR
THE FOUR STSTEMS DESCRIBED AT 55'F OUTSIDE TEMPERATURE

30







I-



10




02 4

UNIT C.OS1 OF ELLCTRICIT -CtNIS PER Kw
O I 2 1 4
COST Or FUEL OIL CENTS PER GALLON
NO.1 DIRECT EXPANSION- ELECTRIC POWEREO SYSTEM
NO.2 INDiROCi EXPANSION-ELECTRIC POWERED SYSTEM
NO 3 ON. TIRED FURNACE 17% EFFICIENCY
N0.4 0 RECT CXPANSON INTERNAL COMBUSTION ENGINE
FIGURE 19
I is the average yearly interest on the investment
CC is the operating cost of the system over a one-year
period
SC Service cost per year.
This equation is based on an expected use of the equipment
for a ten year period.
It can be seen that each individual application must be evalu-
ated separately; however, the results of such a computation on a
hypothetical system are given. The systems were designed to meet
the following conditions:
A building ii Jacksonville. Florida, is to be conditioned to
produce an inside condition not to exceed 80" F dry bulb
and 50 per cent relative humidity when the outside condi-
tions do not exceed 960 F dry bulb and 78z F wet bulb. The
total heat gain of the system including pipe loss and pumping
is computed at 245,000 Btu hr of which 182,(XX) Btu hr is
internal sensible heat. The ratio of sensible to total heat at





peak load is 0.70 and the ratio of internal sensible heat to
total internal heat is .80. The outside air conditions for the
winter cycle design is 250 F dry bulb and the inside tempera-
ture is to be maintained at 700 F dry bulb and 45 per cent
relative humidity. The heat loss from the building is com-
puted at 250,000 Btu/hr. The total air to be circulated is
7,500 cfm of which 1500 cfm is to be fresh air for ventilation
purposes. The winter heat source for systems 2 and 5 and
the summer heat receiver for systems 2, 3 and 5 is a non-
thermal well approximately 50 feet deep.
The operating costs for fuel and electrical energy can be de-
termined from curves similar to those shown in Figures 12, 15
and 18 when used in conjunction with a table such as is shown in
Table 5. This table gives the temperatures which are normally
experienced during each month of the year and at each hour of
the day. Some years vary as much as 30 per cent from these average
normal day data; however, the operating costs on this basis should
be substantially correct for the ten year period of operation.
Table 6 includes six different applications of air conditioning
equipment to satisfy the conditions outlined in the hypothetical
problem stated on page 31. Table 6 shows the general type of
equipment used in each system, the estimated cost of installation
based on a normal year (1937), the summer electrical consumption
in kwh, the winter electrical consumption in kwh and the fuel oil
consumed during the winter cycle and during the summer cycle.
This table does not include the power consumed by the fan deliver-
ing air to the conditioned space. However, it does include all of
the auxiliary electric equipment. It should be noted that system
No. 1 does not have sufficient capacity to heat the space involved
when the outside temperature falls below 400 F and the additional
energy required to provide this auxiliary heat is not included.
System No. 4 requires auxiliary heat when the temperature falls
below 30 F and the energy required for this auxiliary heating is
not included.
From data included in Tables 5 and 6 and curves shown in
Figures 12, 15 and 18, it is possible to draw Figure 20. Figure 20
shows the total owning costs in dollars of each of these systems at
various electric power rates and fuel costs. From Figure 20 the
actual owning costs of systems may be found when the cost of
energy is known. The average cost of electricity from the municipal
electric power plant in Jacksonville is approximately $.025 kwh





TABLE No, 6
DATA ON SYSTEMS DESIGNED FOR AN AIR CONDITIONING INSTALLATION IN JACKSONVILLE, FLORIDA


System No


1. Electric Direct Expansion System usinp
air as heat source 20 HP Compressor her-
mctically sealed. 15 sq. ft. 8 row cvap
orator coil 16.4 sq. ft. 8 row evaporative
condenser.
2. Electric Indirect Expansion System usimn
water at 65" F as heat source. 20 HF
compressor hcrmcricallv scaled 15 sq. ft
5 row water coil water cooled condenser
water chiller
S 3. Electric Direct Expansion System usinm
water at 65 F far condenser water. 20 HF
compressor hermetically scaled 15 sq. ft.
evaporator 8 row, warm air oil fired fur-
nace..
4. Diesel powered Direct Expansion System
Refrigeration Equipment same as for
System No. 1 except open type com-
pressor is used ..
5. Diesel powered Indirect Expansion Sys-
tem refrigeration system same as System
No. 2 except open type compressor is
used..
6. Direct Expansion System similar to Sys-
tem 3 except using evaporative condenser
and nil fired furnace... ..........


Summer
Electric Con-
sumnption kwh





6370


Estimated
Cost of
Installation


Winter
Electric :on-
%unmption kwh





6590




4426


Remarks




Auxiliary heat for temperatures

below 40" F not included




No auxiliary heat required




No. auxiliary heat required


Auxilhary heat required for
temperature below 30 F



No auxiliary heat required


No auxiliary hear required


Fuel Oil

Winter Summer















670



366 450



302 495


670


$7,039.0C


6.246.72




6.640.00



8,019.00



7.246.00


7.140.00










U8 \ \ .
".- r ._ j & o
za" u v


", a, 0 ->
^^5 \O \
d .rJ W Z

z8 ^ 7 4 C


oi '-.

< h j Zd'

__





_o o
e. Ts e i
oi



5 o o u.
Z



3 'S



4Wi


soo v vio





and fuel oil costs approximately $.08 per gallon (1938). With these
data it is possible to draw specific conclusions as to the annual
owning costs of the various systems. If a non-thermal well is avail-
able, the application engineer may choose either of three systems
for heating. The conventional system using a direct expansion
electrically powered compressor with water cooled condenser for
summer air conditioning and an oil fired furnace for winter condi-
tioning has a total annual owning cost of $1,125.00. If an indirect





expansion electrically driven heat pump is used for both summer
and winter conditioning, the annual operating owning cost is
$1,155.00. The Diesel powered heat pump has the lowest annual
owning cost which is $1,040.00.
In case that a non-thermal well is not available, then the ap-
plication engineer might choose from the other three system com-
pared which use evaporative condensers during the summer cycle
and air as a heat source during the winter cycle. The conventional
direct expansion electrically powered compressor and evaporative
condenser for summer conditioning and the oil fired furnace for
winter conditioning has an operating cost, under the conditions
specified, of $1,180.00. The direct expansion electrically powered
compressor and evaporative cuxdeuseL used for both summer and
winter conditioning has an annual owning cost of $1,305.00. The
direct expansion Diesel powered compressors with the evaporative
condenser used for year around conditioning has an annual owning
cost of $1,150. The two latter systems do not include costs for sup-
plementary power if the outside temperature falls below certain
temperatures which have been enumerated previously.
The data used in computing Figure 20 is valid for an installation
located in Jacksonville. Florida. For other locations. results ob-
tained may vary considerably. It will be found that for milder
climates the electrically driven direct and indirect expansion sys-
tems compare more favorably with the conventional summer cool-
ing and winter heating systems. For locations with more extreme
temperatures, it will be found that the Diesel driven direct and
indirect expansion system has a lower comparative owning cost
than is shown in Figure 20. This is true since the cost of producing
heat, as is shown on Figure 19, is lower for the Diesel powered
direct and indirect expansion systems than for the oil fired furnace.

Novel Arrangements of Equipment for Reverse Cycle Refrigeration
Heating: Many so called "tricks" of the refrigeration trade are
employed when equipment is used for reverse cycle heating. Some
of those arrangements used on the compressor include:
(a) Variable compressor speed
(b) Multiple compressors
(c) Valve lifting
(d) Clearance volume variation
(e) Partial cylinder by-pass
(f) Grouping to use more favorable evaporating and condens-
ing temperatures.





Each of these innovations can be used to a particular advantage
on certain types of installations.
Novel arrangements can also be applied to the evaporator and
condenser coil. Five such arrangements are shown in the following
diagrams, which were shown by G. E. Clancy in his article entitled,
"Using the Reverse Cycle for Heating and Conditioning," HEAT-
ING PIPING AND AIR CONDITIONING, June, 1946.11 All of
the arrangements are workable; however, certain arrangements
have particular advantages for certain types of applications.
Figure 13 shows an arrangement whereby the function of the
evaporator and the condenser remain fixed for both the summer
and winter cycles; however, the air stream is diverted. This con-
sistency in the use of the equipment is an advantage because the
equipment may be selected to perform only one function. This is
particularly true in the use of evaporators which sometimes use
distribution headers and refrigerant distributing tubes. This type
of evaporator is inherently not well adapted for use as a condenser
due to the pressure drop through the apparatus. This system is the
type used in the discussion on direct expansion reverse cycle re-
frigeration heating system of this bulletin.
Figure 21 is a similar arrangement to that used in Figure 13
and is classified as semi-regenerative. This system is of use when
the outside air and return air are equal in quantity. The system is
simple and the performance reasonable if the ratio of fresh air
to return air is proper.
The other three systems shown may be used on direct expansion
systems without switching the air flow; however, the refrigerant
flow is switched. Figure 22 is a system of this kind where the
function of the evaporator and condenser are reversed from the
summer to the winter cycle. This arrangement is recommended
only when the evaporator has no special manifold for distribution
of the refrigerant.
The system shown in Figure 23 obviates this dual function re-
quirement; however, the resistance to airflow is increased and con-
sequently the fan horsepower is higher.
Figure 24 is a system similar to that shown in Figure 23 except
that the secondary finned surface for the evaporator condenser is
common. The primary surfaces of the evaporator and condenser
are separated, as separate refrigerant circuits through the coil are
used for each function. This system also uses a liquid refrigerant
sub-cooler on the outside air during the winter cycle. The use of








A- fir 4i A


To
Okf&/4d


Qcashf r#%r
S-8 Ce-A Htn%
*---- CMI/tM Lt~i


IN THIS SCHEME. ON EITHER CYCLE, AIR AT
THE SAME CONDITION IS SUPPLIED TO BOTH
EVAPORATOR AND CONDENSER.
FIGURE 21


Air
Awffm Air
C4daltlWd
fivn0


4~1JJ4]F


IN THIS TYPE OF SYSTEM. THE AIR CIRCUITS
REMAIN UNCHANGED AND THE SPACE RE-
QUIRED IS NOT AS GREAT AS WHEN AIR
SWITCHING IS USED.
FIGURE 22


- /ouT


------ rWh
\ -- *J /'-^C .QHV
llIIClr.
Al cwc,


IN -HIS CASE. THE SWITCHING IS SIMPLIFIED.
FIGURE 23


IN THIS SYSTEM. THE PRIMARY TUBE CIR-
CUITS ARE DUPLICATED. BUT THE SECOND-
ARY SURFACE IS COMMON TO BOTH
CIRCUITS.
FIGURE 24


from
CwWriedmfd


- -


Y r


--


I


+Rr41-1


O r,
JtooLo


rif.





such a device can increase the heat capacity by about 13 per cent.
The self-contained reverse cycle unit manufactured by Drayer
Hanson, Inc., employs the principle shown in Figure 24.
The reader is referred to "A Review of Commercial Heat Pump
Installations" by the Southern Research Institute, Birmingham,
Alabama, for resume and description of reverse cycle heating sys-
tems which have been installed and described in the various trade
magazines. In general, these systems are similar to the ones de-
scribed previously. However, certain adaptations are made on the
particular installations to enable special benefits to be derived.
One particularly interesting feature is that on the installation for
Ohio Power Company at Portsmouth, Ohio. On this installation it
was necessary to provide cooling for certain sections of the building
while heating was used for other sections of the building. There-
fore, the sections to be cooled served as a partial heat source for
the sections to be heated. It can be seen that the design of reverse
cycle heating system depends to a large extent on the ingenuity
of the application or design engineer.
TABLE No. 7
NUMBER OF DAYS WITH MINIMUM TEMPERATURE BELOW 450 F

City Average per Year*
Jacksonville. .............. .... .... 38.15
Miami ... 1.99
Tampa ... .. 14.60
Pensacola.. .... ... 41.94
SAverage number of days with minimum temperature below 45* F per year, based on
Weather Bureau Records for a ten year period.

Heat Pump Installations: There have been several heat pump in-
stallations-some of which are listed in Table 8. This table shows
the location, owner of the system, equipment manufacturers, ap-
proximate date of installation, conditioned volume, type of heat
source, temperature of the heat source, the coefficient of perform-
ance, and supplementary heating equipment. This table is taken
from the "Review of Commercial Heat Pump Installations", by the
Southern Research Institute.'2 More complete information is avail-
able from the literature cited.
Data on the operational cost of reverse cycle heating is quite
meager; however, some data are available on the operation of
several systems.12
The Portsmouth system is installed in a four-story building of
the Ohio Power Company. The dimensions of the building are 104






TABLE No. 8
COMMERCIAL HEAT PUMP INSTALLATIONS


Location


Los Angeles, Calif.
Salem. N.J.
Williamsburg. Va.

Algiers, La.
Whittier, Calif..
Stuhcnvillc, Ohio

Montebello, Calif.
ao Pitman, N.J.. .
< Santa Ana, Calif.
San Bernardino. Calif,.
Emervville. Calif.
New Haven, Conn.

Buenos Aires, Argentina.
Pikcsville, Ky.
Logan, W. Va.

Coshocton. Ohio
Portsmouth. Ohio.
Riverdale, Calif. .
Los Angeles. Calif.
Brilliant. Ohio


Owner


So. Calif. Edison Co.
Atlantic City Electric
Va. Elcc. & Power Co.

La. Power & Light Co
So. Calif. Edison Co.
Ohio Power Co.

So. Calif. Edison Co.
Atlantic City Electric
So. Calif. Edison Co..
So. Calif. Edison Co.
Westinghouse .
United Illuminating Co.

Cine Novcdadcs
Ky. & W. Va. Power Co.
Appalachian Electric
Power Co.
Ohio Power Co.
Ohio Power Co.
So. Sierras Power Co.
Gay Engineering Co.
Ohio Power Co.


Equipm:r.t
Manufacturer

Sturtevant
Gen. Electric
York

York
Westinghouse
Gen. Electric

Frigidaire .
York .
Carrier.
Carrier ..
Westinghouse
Wcstinghouse


Frigidaire .

Carrier
Westinghouse
Gen. Eclctric


Chrysler


Con-
ditioned Heat
Volume Source

3.800.000 OA
76.800 WW
30.800 POA

153.U00 WW
OA
170,000 OA

OA
59,(000 WW
OA
OA
CW OA
WW

700 seats WW
120.000 WW

114.000 WW
WW
OA CW
WW
WW
OA


* Lowest outdoor temperature at whaih heat pump supplies full heating requirements.
OA Outside Air CW City Water
WW Well Water P Process Heat.


Temp.
of Hear
Source

42 WB*
56


77

20*


57


Rated
COP Btu
x 1,000

3.600
260
122

400
10 HP
406

10 tons
162
15 HP
10 HP
238
3.465

325
20 HP


31 KW

1.,000


COP


2.5
3.5


3.5
6.4
3.25

3.81

1.5
3.56
5.0
3.2


3.2


3.6
3

5


Supplementary
Heat
Equipment

EIcc. Resistors
None
50 KW in Res.
Heater
None

15 KW Res
in Stg. Tk.

None



480 KW Res.
in Tank
None
None
Local strip
heater
None
None
None
None
Storage Tank







feet long by 45 feet wide by 451 feet high.1' The system was de-
signed for operation at 00 F outside winter temperature and 950 F
db and 750 F wb outside summer temperatures. The inside design
temperature for winter was 720 F and 301r RH while for summer
the inside conditions were 780 F and 50%, RH. Using these values
the computed maximum winter heat loss was 354,700 Btu hr and
the summer heat gain was 450,000 Btu/hr.
TABLE No. 9
ACTUAL PERFORMANCE DATA FOR PORTSMOUTH SYSTEM

Tests


No. 1 No. 2 No. 3 No. 4

Supply Air to Conditioned Space
Cubic feet per minute.. 18.000 18.000 18.000 18.000
Temperature, degrees Fahrenheit..... 77.1 85.1 78.9 87.2

Outside Air
Cubic feet per minute ..... 2.000 2,000 2.000 2.000
Temperature, degrees Fahrenheit... 32.0 32.0 32.0 32.0

Conditioner Heating Coil
Entering Liquid Temperature, degrees
Fahrcnheit .. ... ... 85.0 97.2 87.5 100.0
Leaving Liquid Temperature, degrees
Fahrenheit ........ ...... 81.0 90.0 83.0 92.0
Gallons per minute 95.0 95.0 95.0 95.0

Liquid Cooler
Entering Liquid Temperature, degrees
Fahrenheit ..... .......... .. 21.2 19.0 28.5 26.0
Leaving Liquid Temperature, degrees
Fahrenheit 18.2 13.8 25.0 20.0
Gallons per minute 95.0 95.0 95.0 95.0

Electric Consumption
Compressor, kilowatts 16.6 32.4 17.8 35.1
Auxiliaries. kilowats .. 9.2 11.1 4.8 4.8
Total kilowatts .. ..... 25.8 43.5 22.6 39.9

Capacity
Refrigeration, BTU/hour........ 140.000 250.000 168.000 284.000
Heat output, BTU/hour .. .. 187.200 342.000 218.000 383.500

Cocfficient of Performance
Using Kilowatt Input to Compress .or. 3.3 3.1 3.6 3.2
Using Total Kilowatt Inputb........... 2.2 2.4 3.0 2.9

Test No.
1-1-25 hp Comp.: 135 psi head pressure-11 psi suction
2-2-25 hp Comp.: 160 psi head pressure-10 psi suction
3-1-25 hp Comp. using water spray-140 psi head pressure-is psi suction
4-2-25 hp Comp. using water spray-170 psi head pressure-12 psi suction
a Kw to conditioner supply fan not included
b 60 per cent of kw input to circulating pumps included as useful work.
40






The system is of the indirect expansion type and is equipped
with modulation capacity valves for 25, 50, 75 and 100'/ capacity.
The performance of the system is given in Table 9.
Another installation for which operating cost data is available
is Cine Novedades, Buenos Aires." This is a theater application for
a 700-seat house. The system employs a water-to-air heat pump.
Well water temperature is 64 F to 680 F at a depth of 75 feet.
The design temperature for Buenos Aires is 230 F winter and 950 F
db and 750 F wb for summer. The summer cooling load is com-
puted to be 579,000 Btu hr. The system is entirely automatic in
its change over from summer to winter operation as this is desirable
on all applications which are subject to relatively high internal heat
variations. This system is also equipped with an economizer fresh
air control which of course reflects a saving in the operating cost
of the installation. Table 10 gives a record of the kilowatt hours
TABLE No. 10

Kilowatt Hours
Hours
Month Compressor and Pump
Fan Heating Cxoling Hearing Cooling Total

August, 1939 230 37.1 585 1.680
September. 1939 220 73.1 1.347 2.462
October. 1939 215 31.0 20.0 393 357 1.655
November, 1939 276 14.9 59.4 245 1.209 2.891
December, 1939. 318 3.7 122.9 56 2.457 4.517
January, 1940 364 1.3 276.0 22 6.669 9,143
February, 1940 401 0.4 229.4 8 5.129 7.653
March, 1940 400 35.5 118.9 602 3.266 6.374
April, 1940. 369 70.6 59.5 1.417 2.031 5.706


of electrical energy used. No record of the degree hours or degree
days for this period of operation is included in the literature.
The Southern California Edison Company, Ltd., Commercial
Office, San Bernardino, California, installation gives an opportunity
for comparison of reverse cycle heating to direct electric heating.
The office was built in 1927 with direct electrical heating but in
1937 the office was remodeled and a 10 HP air to air reverse cycle
heating system was installed. The heating load before and after
remodeling was substantially the same-the only difference being
due to the additional lighting provided, the lighting load was
doubled. Data given in Table 11 gives a comparison of the total
power requirements with each system.







TABLE No. II
OPERATION DATA SOUTHERN CALIFORNIA EDISON COMPANY BUILDING-SAN BERNARDINO, CALIFORNIA


Dar


Av. Power
Input to
Hear Pump


1615

2083

2875

2365

2150

3165

3911

4835

4080

3315

2000

1485


:u from June, 1937 to January. 19
Using Hear Pump

Mean
Monthly Degree Hours
Temp. 65' Critical


53.0

49.9

54.6

61.9

64.6

70.7

77.2

78.2

75.8

66.2

58.7

56.7


8.930

10. 15C

7,740

2.230

300


January....

February..

March

April

May ..

June..

July.

August ..

September ,

October .

November

December


40


Av. Power
for
Lighting


9.200

8.710

8,290

8.420

8,615

7,715

7,797

7,853

7,703

8,380

8,430

8,677


Data from October, 1917 to May. 1937
Using Direct Heating


Av. Power
Input to
Heaters


12,602

11,685

5,833

4,625

2,184

341


435

5,137

10.895


Month


Mean
Monthly
Temp.


50.0

53.4

57.8

60.7

65.0

70.3

77.7

77.9

71.7

65.8

58.8

52,4


_______________ =


Degree Hours
65 Critical


11,160

7,790

5,360

3,100












4,460

9.380


6.180
6.180


Av, Power
for
Lighting


5,156

5,031

4,269

4,832

4,118

3,958

3,604

3,489

3,880

3,678

4,624

5,250





The United Illuminating Company, New Haven, Connecticut,
Administration and Service Building installation also offers a chance
for the comparison of direct electrical heating to heat pump
heating.'", 17, 1 This system utilizes well water at 580 F as a heat
source. Eight 40 HP compressors have their condensers in series so
that each subsequent condenser operates at a higher temperature.
The temperature of the water leaving the last condenser is 1350 F
at an outside temperature of 0 F. This system is also provided
with an auxiliary capacity for direct electric heating, which may
also be used for booster capacity. During January, 1940, the system
operated on direct electrical heat and the electrical consumption
was 208 kwh per degree day. On the following year the reverse
cycle system operated for the month of January at an electrical
consumption of 84 kwh per degree day; Table 12 shows the record
of the operating costs of this system for a three year period.
It will be noted that all comparisons made have been based on
the cost of direct electrical heat, which is to say the least, an ex-
pensive method of heating.

Conclusions: Reverse cycle refrigeration as a method of heating
may be justified by Figure 2 which is a temperature entropy dia-
gram in which the area representing the work done on the com-
pressor is much smaller than the area representing either the cool-
ing or heating effect of the refrigeration system. This figure indi-
cates that the coefficient of performance, or the ratio of heat
removal to the heat input in work done on the compressor, is
greater than one. The figure also indicates that the coefficient of
performance of a refrigeration system as a heater is theoretically
equal to the coefficient of performance of the system as a refrig-
erator plus one. This high coefficient of performance along with
possible savings that might be effected in the original costs of the
summer and winter air conditioning system are the principal
reasons why reverse cycle refrigeration has attracted the interest
of engineers.
Until recently the progress of this type of heating was hindered
by the lack of suitable equipment being manufactured; however,
now this is not an obstacle as the present low pressure type refrig-
eration equipment manufactured for summer air conditioning may
easily be adapted to reverse cycle heating. The machines are small
in size, light in weight and free from vibration, which are the
prerequisites of all air conditioning equipment.
The limitation of reverse cycle refrigeration heating system is






TABLE No. 12
CONSUMPTION OF ELECTRICITY IN KILOWATT-HOURS
Performance of United Illuminating Company, New Haven, Connecticut

1941 1942 1943

Heating Cooling Heating Cooling Heating Cooling

January.. 99.400 95.100 77,700
February..... 80,300 91.000 58,900
March 84.000 63.400 62,900
Apil. 16,900 26,000 44.500
Ma. ..... 8,300 8.800 12,500
June........ 24.800 25,700 29,600
July.. 32.900 37.300 29,600
August......... 28.800 27.600 26,800
September ... 18.100 15.500 12,500
October 4.930 6.800 3.900 16.300 8.200
November.... 37,100 54,300 57,000
December .... 72.100 89,900 84,300
394.700 112.900 426,500 118,800 401,800 119,200

Degree-days. 5.398 5,605 5.897
Kw-hrs. per degree
day ... 73.1 76.0 68.1

OPERATING COSTS OF HEAT PUMP SYSTEM

Cost of
Electricity Operating Mainte-
KW Hrs. at one cent Labor nance Total
per Kw hr.

Heating season,
1941-1942.. 389.600 $ 3.896.00 S 2.539.50 S 704.79 S 7.140.29
Cooling season,
1942 ...... 118.800 1.188.80 2.416.55 214.70 3,819.25
508.400 SS5.084.00 $4.956.05 S 919.49 $10,959.54

Heating season,
1942....... 395,000 3.950.00 2.986.55 856.93 7,703.45
Cooling season,
1943....... 119.200 1.192.00 2.899.16 258.56 4,349.72
514.200 S5.142.00 5.795.71 $1.115.46 S12.053.17


that of reduced


capacity when working as a heater under certain


conditions. The capacity of any refrigeration system is reduced
with a reduction of refrigerant evaporating temperature or an in-
crease in condensing temperature. Therefore, for this type of ap-
plication, a heat source must be obtained at a sufficiently high
temperature to maintain the desired heat capacity from the

44






machine, and a heat discharge at low enough temperatures to pre-
vent the condensing pressures from becoming excessive. The oper-
ating costs and annual owning cost must always be considered if
an intelligent application of equipment is to be made.
The heat source using direct expansion may be either outside
air or water sprayed over the evaporator coil. Difficulty has been
found in keeping the refrigerant evaporating temperature above
320 F using a direct expansion system. The evaporating tempera-
ture can be raised by the addition of more evaporator coil surface;
however, twice the surface required for summer operation has been
found to be the usual limit of increase in the amount of surface
that is economically feasible to use.
Table 7 gives the average number of days for which a minimum
temperature below 45" F is experienced in each of four Florida
cities. A mixed air temperature of 40" is considered the minimum
temperature for air to enter the evaporator of a direct expansion
reverse cycle heating system in order to prevent the formation of
ice on the evaporator. Therefore, it is seen that some auxiliary
method of heating would probably be necessary to serve the con-
ditioned area for the number of days indicated in Table 7. This
auxiliary heat might be the gas or electric strip. It can be seen that
Miami, Florida has less than two days per year in which the tem-
perature falls below 45" F. Therefore, the direct expansion reverse
cycle refrigeration heating system shows considerable promise for
applications in the Miami area.
The performance of the direct expansion reverse cycle heating
system is given on Figure 15, the heating capacity being plotted
as a function of air temperature entering the evaporator. Since it
is possible to mix exhaust air from the space (the amount not to
exceed the amount of fresh air introduced to the space) with the
outside air and introduce this mixture to the evaporator, higher air
temperatures entering the evaporator may be obtained than when
using 100 per cent outside air. The use of such a feature increases
the capacity of this type of system.
The cost for heating per therm for four types of winter heating
are given on Figure 19. It can be seen that the cost of heating with
electrically powered reverse cycle heating systems is greater than
the energy cost for heating with fuel oil. The cost of fuel oil for
heating with internal combustion engine powered reverse cycle
refrigeration units is approximately one-half of that required for
heating with oil furnaces.








Figure 20 compares the annual cost of owning six year-round
air conditioning systems for an application in Jacksonville, Florida.
It can be seen that the total owning cost of the Diesel indirect
expansion system operating on $.08 per gallon fuel oil is lower
than any of the other types of heating systems compared, if the
electric rate is greater than $.011 per kwh. The direct expansion
systems 1 and 4 shown on this figure involves the use of air cooled
condensers and therefore, the initial cost is higher than it would
be if water cooled condensers were used with direct expansion
evaporators. It should be realized that the data included in Figure
20 is valid for the particular application which is located in Jack-
sonville, Florida. In general, direct expansion systems will compare
more favorably than indicated in locations of milder climates such
as Miami, Florida. Diesel powered systems will compare more
favorably than is indicated by Figure 20 in locations with more
extreme climates. Diesel direct expansion units may be found to
be impractical if the climatic conditions are too extreme and for
this reason, each application should be considered independently.
The climatic data used for estimating the operating cost is based
on an average condition experienced during the ten year period
of 1927 through 1936. It will be found that in certain years varia-
tions from normal or average conditions may be as high as 30 per
cent. However, the total operating costs over a ten year period
should be substantially the same as those indicated by these solu-
tions. The cost used in comparison of investment costs and interest
on the various systems are based on installation prices during the
year of 1937, which is considered to be normal.








GLOSSARY OF ABBREVIATIONS
CP h Coefficient of performance as a heater
CP Coefficient of performance as a refrigerator
gph Gallons per hour
gpm Gallons per minute
H Enthalpy
I Average yearly interest on investment
n Exponent for polytropic compression
OC Annual operating cost of air conditioning system (fuel or
electricity)
P Pressure-lb. per sq. in absolute
Q QrorQc
Qc Heat dissipated at the condenser-Btu/hr.
Q, Heat absorbed or refrigeration effect at the evaporator Btu/hr.
RD Refrigerant discharge line
RL Refrigerant liquid line
RS Refrigerant suction line
TC Total cost of installation
T, Refrigerant condensing temperature in degrees F absolute
T e Refrigerant evaporating temperature in degrees F absolute
tm Log mean temperature difference
U Overall coefficient of heat transfer-Btu/F sq. ft.
v Specific volume-cu. ft. per lb.






BIBLIOGRAPHY

1. "Heating and Cooling Buildings by Means of Currents of Air," GLAS-
GOW PHIL. SOCIETY PROCEEDING, Vol. 3, December, 1852.
2. "The Heat Pump -An Economical Method of Producing Low Grade Heat
from Electricity,' by T. G. N. Haldane, ELECTRIC REVIEW, Vol. 105,
pp 1161-1162, December 27, 1929.
3. "Engineering Thermndynamics," N. C. Ebaugh. D. Van Nostrand Co.
4. "Westinghouse Air Conditioning and Industrial Refrigeration Equip-
ment," page CLS 640/850-BI.
5. "A Survey of Heat Pump Sources," E. N. Kemler, Southern Research
Institute, Birmingham, Alabama, September 5, 1946.
6. U. S. Geological Survey, Water Supply Paper 520.

7. AMERICAN SOCIETY OF HEATING & VENTILATING ENGINEERS'
GUIDE FOR 1946.
8. "The Economic Possibilities of Reversed Cycle Refrigeration for Heating
in Florida," S. P. Goethe-Thesis, University of Florida, 1938.

9. "Climatic Data for the Design and Operation of Air Conditioning Systems
in Florida," N. C. Ebaugh and S. 1'. Goethe, Florida Engineering & In-
dustrial Experiment Station Bulletin No. 5, 1939.

10. "How to Solve Heat Transfer Problems," William Goodman, HEATING,
PIPING AND AIR CONDITIONING Magazine, Vol. 9, Nos. 1, 2, 3-
January, February, March, 1937.
11. "Using the Reverse Cycle for Heating, Conditioning," G. E. Clancy,
HEATING, PIPING AND AIR CONDITIONING, June, 1946.
12. "Review of Commercial Heat Pump Installations," Sabert Oglesby, Jr.,
Southern Research Institute, Birmingham, Alabama.
13. "Description and Performance of Two Heat Pump Air Conditioning
Systems," P. Sporn and E. R. Ambrose, HEATING, PIPING AND AIR
CONDITIONING, June, 1944.
14. "South American Way on this Buenos Aires Job was to Use a Heat
Pump," M. A. Ramsey, HEATING, PIPING AND AIR CONDITIONING,
March, 1941.
15. Electric Heating Bulletin No. 52-Southern Edison Company, Ltd., Los
Angeles, California-January 30, 1940.
16. "Heating by Reversed Refrigeration," Lawless, A. J., HEATING, PIPING
AND AIR CONDITIONING-Aug. 1940, Sept. 1940.
17. "Fvperience with a Reverse Cycle Heating System," POWER, August,
1941.

18. "Heating by Reverse Refrigeration," R. D. Hutchue, ELECTRICAL
ENGINEERING-November, 1942.





PUBLICATIONS OF THE FLORIDA
ENGINEERING AND INDUSTRIAL EXPERIMENT STATION
As long as the supply is adequate, copies of available publications are free for
general distribution. Address all requests to: The Director, Florida Engineeing
and Industrial Experiment Station, University of Florida, Gainesville, Florida.
BULLETIN SERIES
No. 1 "The Mapping Situation in Florida," by William L. Sawyer.
No. 2 "The Electrical Industry in Florida," by John W. Wilson.
No. 8 "The Locating of Tropical Stoms by Means of Associated Stati," by
Joseph Well and Wayne Mason.
No. 4 Study of Beach Conditions at Daytona Beach, Florida, and Vicianty,"
by W. W. Fiaren.
No. 5 "Climatic Data for the Design and Operation of Air Conditioning Sys-
tems in Florida," by N. C. Ebaugh and S. P. Goethe.
No. 6 "On Static Emanating from Six Tropical Storms and its use in Locating
the Position of the Disturbance," by S. P. Sashoff and Joseph Wel.
No. 7 "Lime Rock Concrete-Part I" by Harry H. Houston and Balph A.
Morgen.
No. 8 "An Industrial Survey of Hides and Skins in Florida," by William D.
May.
No. 9 "Studies on Intermittent Sand Filtration-Part L" by D. L. Emeson, Jr.
No. 10 "Florida Spray Gun for Pine Tree Gum Flow Stimulation," by Norman
Bourke and Keith W. Dorman.
No. 11 "Developments of Ceramic Compositions Suitable for the Production of
Porcelain Type Artware" by B. W. Thorngate.
No. 12 "Mold and Mildew Control for Industry and the Home" by S. S. Block.
No. 18 "Engineering and Industrial Research at the University of Florida."
TECHNICAL PAPER SERIES
No. 1 Heats of Solution of the System Sulfur Trioide and Water, by Ralph
A. Morgen.
No. 2 The Useful Life of Pyro-Meta and Tetrapbosphate, by Ralph A. Margua
and Robert L Swoope.
No. S Florida Lime Rock as an Admixture in Mortar and Concrete, by Harry
H. Houston and Ralph A. Morgen.
No. 4 Country Hides and Skins, by William D. May.
No. 5 Empirical Correction for Compressibility Factor and Activity Coefficient
Curves, by R. A. Morgen and J. H. Childs.
No. 6 Crate Closing Device, by William T. Tiffin.
No 7 The System Sodium Acetate-Sodium Hydroxide-Water, by R. A. Mrgen
and R. D. Walker, Jr.
No. 8 Patent Policies for Sponsored Research, by Ralph A. Morgan.
No. 9 Conservation of Municipal Water Supplies in Air-Conditioning Systems,
by N. C. Enough.
No. 10 Florida Scrub ak-New Source of Vegetable Tannin, H. N. Calerwood
and W. D. May.
No. 11 Protein Feed from Sulfite Waste Liquor, Ralph A. Morgen and Robert D.
Walker, Jr.
No. 12 Effect of Moisture on Thermal Conductivity of Limerock Cancrete, by
Mack Tyner.
No. 18 Insect Tests of Wire Screening Effectiveness, S. S. Block.
No. 14 Properties of Limerock Concrete, Mack Tyner.
No. 15 Scrub Oak as a Potential Replacement for Chestnut, by H. N. Caldm
wood and W. D. May.




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