fflSERDP
EPA-600/R-97-058
June 1997
Envrjrmertal R«<«Ch
«f*1 3flv©
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NOTICE
This document has been reviewed in accordance with U.S. Environmental
Protection Agency policy and approved for publication. Mention of trade
names or commercial products does not constitute endorsement or
recommendation for use.
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FOREWORD
The U.S. Environmental Protection Agency is charged by Congress with pro-
tecting the Nation's land, air, and water resources. Under a mandate of national
environmental laws, the Agency strives to formulate and implement actions lead-
ing to a compatible balance between human activities and the ability of natural
systems to support and nurture life. To meet this mandate, EPA's research
program is providing data and technical support for solving environmental pro-
blems today and building a science knowledge base necessary to manage our eco-
logical resources wisely, understand how pollutants affect our health, and pre-
vent or reduce environmental risks in the future.
The National Risk Management Research Laboratory is the Agency's center for
investigation of technological and management approaches for reducing risks
from threats to human health and the environment. The focus of the Laboratory's
research program is on methods for the prevention and control of pollution to air,
land, water, and subsurface resources; protection of water quality in public water
systems; remediation of contaminated sites and groundwater; and prevention and
control of indoor air pollution. The goal of this research effort is to catalyze
development and implementation of innovative, cost-effective environmental
technologies; develop scientific and engineering information needed by EPA to
support regulatory and policy decisions; and provide technical support and infor-
mation transfer to ensure effective implementation of environmental regulations
and strategies.
This publication has been produced as part of the Laboratory's strategic long-
term research plan. It is published and made available by EPA's Office of Re-
search and Development to assist the user community and to link researchers
with their clients.
E. Timothy Oppelt, Director
National Risk Management Research Laboratory
i i i
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ABSTRACT
In compliance with the Montreal Protocol and Department of Defense directives, alternatives to
refrigerant CFC-114 are being investigated by the U.S. Navy and the U.S. Environmental Protection Agency for
use in shipboard chillers. Refrigerant HFC-236ea has emerged as a candidate for drop-in replacement.
A computer model was developed for comparing these two refrigerants in a simulated 440-kilowatt
centrifugal chiller system. Equations for modeling each system component were developed and solved using the
Newton-Raphson method for multiple equations and unknowns. Correlations were developed for CFC-114 and
HFC-236ea boiling and condensing coefficients taken at the Iowa State Heat Transfer Test Facility. The model
was tested for a range of inlet condenser water temperatures and evaporator loads. The results are presented and
compared with data provided by the Naval Surface Warfare Center in Annapolis, MD.
The experimental data provided by the Naval Surface Warfare Center sufficiently validate the model, and
the simulation model predicts that HFC-236ea would perform favorably as a drop-in substitute for CFC-114.
Several recommendations are discussed which may further improve the performance of HFC-236ea in
Navy chillers. Recommendations include adjusting the load of the evaporator to achieve positive gage pressure,
use of a purge device, use of a variable speed compressor, further testing with azeotropic mixtures, and use of high
performance tubes in the heat exchangers.
This report was submitted in fulfillment of Cooperative Agreement No. CR 820755-01-4 by the
Engineering Research Institute, College of Engineering, Iowa State University, Ames, IA, under the sponsorship of
the United States Environmental Protection Agency with funding from the Strategic Environmental Research and
Development Program* (SERDP). This work covers the period from October 1, 1992 to May 3, 1995, and the
work was completed as of May 3, 1995.
(*) A joint program of the Department of Defense, the Department of Energy, and the Environmental Protection
Agency.
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TECHNICAL REPORT DATA
(Please read Instructions on the reverse before comp,
1. REPORT NO. 2.
EPA-600/R-97-058


4. TITLE AND SUBTITLE
Comparison of CFC-114 and HFC-236ea Performance
in Shipboard Vapor Compression Systems
5. REPORT DATE
June 1997
6. PERFORMING ORGANIZATION CODE
7. AUTHOR(S)
Daniel T. Ray, M, B. Pate, and H. N. Shapiro
8. PERFORMING ORGANIZATION REPORT NO.
9. PERFORMING ORGANIZATION NAME AND ADDRESS
Iowa State University
2088 IT. M. Black Engineering Building
Ames, Iowa 50011-2160
10. PROGRAM ELEMENT NO.
11. CONTRACT/GRANT NO.
CR820755-01-4
12. SPONSORING AGENCY NAME ANO ADDRESS
EPA, Office of Research and Development
Air Pollution Prevention and Control Division
Research Triangle Park, NC 27711
13. TYPE OF REPORT AND PERIOO COVERED
Final; 10/92-5/95
14. SPONSORING AGENCY CODE
EPA/600/13
is.supplementary notes project officer is Theodore G. Brna, Mail Drop 04, 919/
541-2683.
16. abstract 'X"hc report gives results of a comparison of the performance of two refri-
gerants—1,1,1, 2, 3, 3-hexafluoropropane (HFC-236ea) and 1, 2-dichloro-tetrafluoro-
ethane (CFC-114)--in shipboard vapor compression refrigeration systems. (NOTE:
In compliance with the Montreal Protocol and Department of Defense directives, al-
ternatives to CFC-114 are being investigated by the U. S. Navy and the U.S. EPA for
use in shipboard chillers. IIFC-236ea has emerged as a candidate for drop-in re-
placement. ) A computer model was developed for comparing the two refrigerants
in a simulated 440-kW centrifugal chiller system. Equations for modeling each sys-
tem component were developed and solved using the Newton-Raphson method for mul-
tiple equations and unknowns. Correlations were developed for CFC-114 and HFC-
236ea boiling and condensing coefficients taken at the Iowa State Heat Transfer Test
Facility. The model was tested with data provided by the Naval Surface Warfare
Center (NSWC) in Annapolis, MD. The experimental data provided by the NSWC suf-
ficiently validate the model, and the simulation jnodel predicts that HFC~236ea
would perform favorably as a drop-in substitute for CFC-114. Several recommenda-
tions are discussed which may further improve the performance of IIFC-236ea in
Navy chillers.
17. KEY WORDS AND DOCUMENT ANALYSIS
a. DESCRIPTORS
b.IDENTIFIERS/OPEN ENDED TERMS
c. cosati Field/Group
Pollution
Refrigerants
Performance
Halohydrocarbons
Mathematical Models
Pollution Control
Stationary Sources
Shipboard Chillers
Vapor Compression
Chlorofluorocarbons
Hydrofluorocarbons
13 B
13A
14G
07C
12 A
18. DISTRIBUTION STATEMENT
Release to Public
19. SECURITY CLASS (This Report)
Unclassified
21. NO. OF PAGES
85
20. SECURITY CLASS (This page)
Unclassified
22. PRICE
EPA Form 2220-1 (9-73)
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CONTENTS
ABSTRACT	i V
FIGURES	VI1
TABLES	 Vili
NOMENCLATURE	 ix
ACKNOWLEDGEMENT	 xi
1.	INTRODUCTION	1
Ozone Depletion	1
Global Warming	3
Montreal Protocol	4
Refrigeration Industry's Response	5
CFC-114 and the U.S. Navy	6
Fiuorinated Hydrocarbons	7
Summary	7
2.	CONCLUSIONS	9
3.	RECOMMENDATIONS	10
4.	REVIEW OF RELATED LITERATURE	11
5.	VAPOR-COMPRESSION CYCLE	13
Condenser	15
Evaporator	16
Compressor	17
Expansion Device	18
6.	MODEL DESCRIPTION	19
Properties	20
Program Description	20
Thermodynamic Analysis	22
Heat Transfer Analysis	25
Numerical Procedure		29
7.	RESULTS AND DISCUSSION	30
Comparison of Model and Experimental Results	30
Comparison of CFC-114 and HFC-236ea Performance	37
Summary	55
8.	REFERENCES	56
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Contents (continued)
APPENDIX A. COMPUTER PROGRAM	58
APPENDIX B. NSWC AC PLANT INSTRUMENTATION SCHEMATIC	73
APPENDIX C. SAMPLE NSWC DATA	74
APPENDIX D. CHILLER INFORMATION PROVIDED BY THE NSWC	75
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FIGURES
Number
5.1: Components of a vapor-compression refrigeration system	13
5.2: T-s diagram of a typical vapor-compression refrigeration cycle	14
5.3: Diagram of a typical shell-and-tube condenser	16
5.4: Diagram of a typical shell-and-tube flooded evaporator	17
6.1: Program flow diagram	21
6.2: P-h diagram for CFC-114 and HFC-236ea	23
7.1: Comparison of modeled and measured compressor power	31
7.2: Comparison of modeled and measured coefficient of performance	32
7.3: Comparison of modeled and measured condenser temperature	34
7.4: Comparison of modeled and measured condenser capacity	35
7.5: Comparison of modeled and measured leaving condenser water temperature	36
7.6: Comparison of modeled and measured refrigerant saturation temperatures in the evaporator	36
7.7: Comparison of modeled and measured refrigerant flow rate	37
7 8: Dependence of compressor power requirement on entering condenser water temperature	39
7.9: Dependence of refrigerating performance on entering condenser water temperature	40
7 .10: Dependence of refrigerating efficiency on entering condenser water temperature	41
7,11: Dependence of condenser temperature on entering condenser water temperature	42
7.12: Dependence of evaporator temperature on entering condenser water temperature	43
7.13: Dependence of evaporator capacity on entering condenser water temperature	44
7.14: Dependence of condenser capacity on entering condenser water temperature	44
7 15: Dependence of refrigerant mass flow rate on entering condenser water temperature	45
7.16: Dependence of compressor power for CFC-114 and HFC-236ea on entering evaporator
water temperature	46
7.17: Dependence of refrigerating performance on entering evaporator water temperature	47
7.18: Dependence of refrigerating efficiency on entering evaporator water temperature	48
7.19: Dependence of compressor power input on leaving evaporator water temperature	49
7.20: Dependence of refrigerating performance on leaving evaporator water temperature	50
7.21: Dependence of refrigerating efficiency on leaving evaporator water temperature	51
7.22: Dependence of power on evaporator water flow rate	52
7.23: Dependence of refrigerating performance on evaporator water flow rate	52
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Figures (continued)
7.24: Dependence of refrigerating efficiency on evaporator water flow rate	53
7.25: Dependence of power on condenser water flow rate	54
7.26: Dependence of refrigerating performance on condenser water flow rate	54
7.27: Dependence of refrigerating efficiency on condenser water flow rate	55
TABLES
Table 1.1: Refrigerant comparisons		3
Table 4 1: Comparison of models found in the literature	12
Table 6.1: Simulated design conditions	19
Table 7.1. Effect of entering condenser water temperature on measured and modeled COP for HFC 236ea . 33
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NOMENCLATURE
Afc
total tube flow area of condenser (m~)
Afe
total tube flow area of evaporator (m^)
Ai,c
total condenser tube inside surface area (m2)
Ai,e ~~
total evaporator tube inside surface area (m2)
Ao,c ~~
total outside surface area of condenser (m2)
Ao.e
total outside surface area of evaporator (m*-)
Aa.c
average outside surface area of condenser per length (m2/m)
Aoe
average outside surface area of evaporator per length (m2/m)
^Pchw
specific heat of chilled water (ld/kg-°C)
CPsw
specific heat of sea water (kJ/kg-°C)
d;,c ~
inside diameter of a condenser tube (m)
dj,e ~
inside diameter of an evaporator tube (m)
h
enthalpy (kJ/kg)
c ~
condenser water side heat transfer coefficient (kW/m2-°C)
\e
evaporator water side heal transfer coefficient (kW/m2-°C)
^o. c ~
condenser refrigerant side heat transfer coefficient (kW/m2-uC)
ho,e ~~
evaporator refrigerant side heat transfer coefficient (kW/m2-°C)
kchw
thermal conductivity of chilled water (kW/m-°(?)
^sw =
thermal conductivity of sea water (kW/m-°C)
l4f-c
effective tube length per pass in condenser (m/tube)
II
effective tube length per pass in evaporator (m/tube)
mchw
mass flow rate of dulled water (kg/sec)
mr
mass flow rate of refrigerant (kg/min)
msw
mass flow rate of condensing sea water (kg/sec)
n =
polytropic exponent
Ne
number of tubes in evaporator
Nc
number of tubes in condenser
P
absolute pressure (kPa)
Frchw ~~
average Prandtl number of chilled water in evaporator
Frsn>
average Prandtl number of sea water in condenser
Qc
heat transferred in condenser (kW)
Qe
heat transferred in evaporator (kW)
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Nomenclature (continued)
He"
2
heat flux in the condenser (kW/m )
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ACKNOWLEDGEMENT
This research project was funded by the United States Environmental Protection Agency, National Risk
Management Research Laboratory [formerly the Air and Energy Engineering Research Laboratory (AEERL)],
through the support of the DOD/DOE/EPA Strategic Environmental Research and Development Program
(SERDP). The authors wish to thank the EPA staff for their suggestions and support.
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CHAPTER 1. INTRODUCTION
Fully halogenated chlorofluorocarbons (CFCs) are manufactured chemicals with properties that make
them useful for such applications as aerosol propellants, foam blowing agents, solvents, and refrigerants for
automotive, residential, and commercial applications. Introduced in the U.S. in the 1930's, the use of CFCs grew
steadily after World War II and today they play a prominent role in human lifestyle and comfort CFCs became
popular in part because they were chemically stable, non-flammable, and non-toxic. Ironically, the chemical
stability of CFCs is the cause for their present perceived threat to the environment. Scientific evidence suggests
that the harmful alterations of the earth's atmosphere occurring from the use of CFCs are of regional and global
proportions. As early as 1974, concerns about the potential harmful environmental effects associated with the use
of CFCs were raised when it was suggested that the chlorine from these compounds could efficiently destroy
stratospheric ozone [1], Additionally, there is a growing consensus among scientists that CFCs may contribute to
global warming [2], In the 1970s, regulatory action banning selected, non-essential CFC compounds used as
aerosol propellants temporarily decreased the release of chlorine-containing compounds into the atmosphere.
However, the increased use of CFCs, in part by newly industrialized and lesser developed countries, has resulted in
the need for stronger control measures. Moreover, hydrogenated chlorofluorocarbons (HCFCs), once thought to be
an acceptable replacement for CFCs. have also been implicated as potentially harmful to the environment.
Significant steps have been taken to eliminate CFC and HCFC consumption including restrictive
legislation, the development of alternative refrigerants, and the pursuit of new technologies. Policy makers,
industry leaders, and researchers worldwide have recognized the need for continued efforts to understand the
potential long range impacts that the use of these chemical compounds and their replacements may have on the
environment, lifestyle, and economy: on environment, because climate changes and health problems could be
significant, on lifestyle, because humans have come to appreciate and demand the comforts of air-conditioning and
refrigeration; and on economy because of the potential need to redesign and replace billions of dollars of existing
equipment and chemicals (refrigerants).
OZONE DEPLETION
Ozone exists naturally in the upper stratosphere and is a primary absorber of ultraviolet radiation. Ozone
concentration determines the amount of ultraviolet radiation reaching the earth's surface. Ozone molecules are
broken apart by high energy ultraviolet radiation from the sun and rapidly re-form to maintain a relatively stable
level of ozone in the stratosphere. The presence of chlorine in the stratosphere disrupts this natural balance.
Chlorine from natural sources is washed out of the air by rain before it can migrate to the stratosphere.
Methyl chloride, given off by ocean plankton, appears to be an exception; however, measurements show it accounts
for only one sixth of the chlorine in the stratosphere [3J. Synthetic compounds such as chlorofluorocarbons, on the
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other hand, make their way to the stratosphere and disrupt the natural balance of ozone by a series of rapid
reactions. Intense ultraviolet light in the stratosphere splits apart the CFC molecule and releases a chlorine atom.
The free chlorine radical reacts with ozone, breaking it into an ordinary oxygen molecule and forming a chlorine
monoxide molecule. Chlorine monoxide can combine with a single oxygen atom to form a second oxygen
molecule. The chlorine atom, freed in this reaction, can then repeat its ozone-destroying cycle a hundred thousand
times before being converted to a less reactive form that is eventually removed from the stratosphere by natural
processes [3J. Models that simulate this chain suggest that CFCs have long atmospheric lifetimes ranging from
decades to centuries [1], Because of this, it is estimated that concentrations of chlorine in the stratosphere will
continue to increase for some period even after CFC emissions cease. The consequences of increased ultraviolet
radiation reaching the earth's surface include negative impacts on human health and possible changes in aquatic
and terrestrial ecosystems, the total ramifications of which are largely uncertain.
HCFCs retain many of the desirable properties of CFCs; however, as a result of the hydrogen in their
molecular structure, they have much shorter lifetimes in the atmosphere. Consequently, their potential effects on
ozone and the climate are significantly reduced compared with the compounds they replace. However, with
increased use, they could significantly contribute to environmental problems. Hydrogenated fluorocarbons (HFCs),
another alternative to CFCs, have no chlorine in their structure and consequently provide no contribution to ozone,
destruction.
Predicting trends in the ozone depletion rate is very difficult because local ozone concentrations vary with
altitude, latitude, temperature, and seasonal changes; they are also affected by natural processes such as air
currents. A scale has been developed, based on complex models, to attempt to compare the ozone depiction
potential (ODP) of various compounds against CFC-12, which by definition has an ODP of one. For example,
CFC-114 has an ODP of 0.7 and HFC-236ea has an ODP of 0 Table 1.1 includes a comparison of ODPs for
refrigerants of interest in this study.
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Table 1.1: Refrigerant comparisons a
\
Cur
CFfc m
rent
rfgerante
Potential
Refrigerant
Designation
CFC-11
CFC-114
HCFC-124
FC-318
E-134
HFC-236ea
Chemical
Formula
CCI3F
c2ci2f 4
c2hcif4
C4I=8
chf2-o-
chf2

Evaporator
Pressure (kPa)
48.69
103.38
192.07
152.62
94.97
94 41
Condenser
Pressure (kPa)
161.45
311.72
557.45
457.72
322.55
313.93
Flowrate
(m3/min/ton)
0.4517
0.2636
0.1433
0.1960
0.2413

Power (kW/ton)
0.463
0.518
0.499
0.540
0.488

Discharge
Superheat (nC) or
Qualify (%)
-10.5 °C
96%
99%
90%
-16.4 °C

Speed of Sound
at Suction (m/sec)
134.7
115.2
128.3
120.1
141.7
122.8
Ozone Depletion
Potential
1.0
0.7
0.02
0
0
0
Global Warming
Potential
1.0
3.7
0.07
>1
low
low
Atmospheric
Lifetime (yr)
75
200
8
>100
short
1.2
Acute Toxicity
(1 = hlqh, 6 = low)
5
6
67
6 (est.)
6 (est.)
6 (est.)
Flammable
no
no
no
no
no
no
aProperties are from NIST REFPROP computer program version 4.01.
Data are based on an evaporating temperature of 4.8°C, a condensing
temperature of 38.1°C, and compressor and motor efficiencies of 100%.
GLOBAL WARMING
A natural energy balance exists in the earth and its atmosphere between absorption of solar radiation 3tid
emission of infrared radiation to space. Greenhouse gases in the atmosphere are relatively inefficient absorbers of
incoming short wavelength energy but strong absorbers of outgoing infrared radiation Greenhouse gases trap heat
that would otherwise radiate from earth leaving the planet with a much colder average surface temperature than
the planet's current average surface temperature of 288 K [2] The concentrations of greenhouse gases in the
troposphere determine the net trapping of heat in the atmosphere. To maintain a global energy balance, the likely
effect of an increase in greenhouse gas concentrations is a change in atmospheric and surface temperature
Interestingly, in addition to absorbing ultraviolet radiation, ozone is also a greenhouse gas An observed decrease
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in stratospheric ozone over the last decade suggests a global cooling tendency. This cooling tendency, when
globally averaged, is comparable in magnitude and opposite in sign to the estimated warming from increased CFC
concentrations in the troposphere [2],
Greenhouse gases include water vapor, carbon dioxide, nitrous oxide, CFCs, HCFCs, HFCs, methane, and
tropospheric ozone; each is an absorber in specific bands within the infrared spectrum. CFCs, IiCFCs and HFCs
happen to absorb energy in the wavelength window of 7 to 13 fim where the primary absorbers—carbon dioxide and
water vapor~are weak radiation absorbers. Absorption in this region allows gases with much smaller atmospheric
concentrations than carbon dioxide and water vapor to exert significant radiative forcing on climate resulting in
linear increases in infrared absorption with increasing atmospheric concentration. In contrast, carbon dioxide,
having a large atmospheric concentration, already absorbs essentially all of the radiation in the central cores of its
absorption lines and will increase infrared absorbtion only slightly with further increased concentration [2], In
other words, comparable increases in the concentration of different greenhouse gases may have vastly different
greenhouse effects.
Concern about potential global wanning exists because there is a wide range of possible negative effects
that include changes in sea level and changes in local climates—the consequences of which are not well understood.
There are large uncertainties in predicting greenhouse effects. For example, a major source of uncertainty comes
from a poor understanding of cloud dynamics. A scale has been developed to rate the global warming potential
(GWP) of various refrigerants relative to the effects of carbon dioxide. Table 1.1 includes a comparison of the
G WP for refrigerants of interest in this study.
While refrigerants escaping into the troposphere have a direct effect on global warming, fossil fuel energy
consumed by refrigerant systems provides an additional indirect contribution to global warming by adding carbon
dioxide to the atmosphere. Thus, system efficiency is an important consideration in determining the suitability of
replacement refrigerants.
MONTREAL PROTOCOL
In September 1987, delegates to the United Nations Environment Program (TJNEP) signed the Montreal
Protocol for Substances that Deplete the Ozone Layer and, thereby, agreed to limit production of CFCs and haJons.
Spurred by alarming decreases in stratospheric ozone concentrations in the Antarctic region, UNEP delegates
amended the Protocol in 1990 and again in 1992 to broaden the scope of substances covered and to accelerate their
phase-out. Similar restrictions were enacted by the United States Congress in the 1990 Clean Air Act
Amendments. The Department of Defense (DOD) and the Secretary of the Navy have also issued directives for the
Navy's compliance with these policies [4], Currently, CFCs are scheduled to be phased out of production
completely by the end of 1995.
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The Environmental Protection Agency has also finalized an accelerated schedule to phase out the
production of HCFCs. The latest schedule is as follows:
•	By 2003: Ban on production of HCFC-141b
•	By 2010: Production frozen at baseline levels for HCFC-22 and HCFC-142b
•	By 2010: Ban on use of virgin chemical unless used as a feed stock or refrigerant in appliances manufactured
prior to Jan 1, 2010, for HCFC-22 and HCFC-142b
•	By 2015: Production freeze at baseline levels for all other HCFCs
•	By 2020: Ban on use of virgin chemical unless used as a feed stock or refrigerant in appliances manufactured
prior to Jan 1, 2020, for all other HCFCs
•	By 2020: Ban on production of HCFC-22 and HCFC-142b
•	By 2030: Ban on production of all other HCFCs [5]
While these measures were initialed in response to evidence of ozone destruction, a similar movement is
underway that may lead to an international protocol regarding the use of substances that contribute to global
warming This could affect the future use of HFCs that currently have no restrictions placed on them.
REFRIGERATION INDUSTRY'S RESPONSE
The refrigeration industry has responded to restrictions on the production of CFCs and HCFCs by
developing new environmentally safe refrigerants and refrigerant mixtures with similar thermodynamic and heat
transfer characteristics. These alternatives will serve as near term substitutes for existing air-conditioning and
refrigeration equipment with remaining useful life. Additionally, new emphasis is being placed on research and
development of cooling systems based on emerging new technologies, including alternatives to typical vapor-
compression systems. Thus, a two-fold challenge to the refrigeration industry entails replacing CFC refrigerants in
existing equipment in the near term, and designing efficient, environmentally safe cooling systems for the future.
The urgency of the situation has been emphasized by the series of accelerations to the original phase-out schedule
put forth in the Montreal Protocol.
The success of HFC-134a as a replacement for CFC-12 provides an example of the rapid progress that has
been made toward replacing refrigerants targeted for elimination, but also illustrates some of the problems
encountered along the way. HFC-134a has emerged as a near drop-in replacement for existing CFC-12 systems.
Its success has resulted in commercial availability in new products such as new air conditioners for cars. A "drop-
in" replacement implies that only minimal and low-cost changes will need to be made in order for the refrigeration
system to accept a new refrigerant; this poses several challenges in finding an appropriate alternative refrigerant.
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Similar thermodynamic and heat transfer properties are desired which will minimize the changes in efficiency,
power consumption, size, volume, and operating pressures of the original system. Similar refrigerant properties
will also minimize the need to make costly modifications to the system components such as the heat exchangers
and compressors. Material compatibility is a concern because different refrigerants may not be compatible with
seals, gaskets, diaphragms, and flexible hoses in the original system. This is also of concern when changing
lubricants. Finally, the conversion process itself may be restricted by the nature and importance of the application.
For example, a supermarket or a hospital may have to plan carefully so as not to interrupt critical cooling while
conversions to a new refrigerant are being made.
While the intense effort to replace CFC-12 has been successful, there is a need to find replacements for
other CFCs that are not as widely used, yet are included in the world wide CFC ban. One such refrigerant that
needs a suitable replacement is CFC-114, whose characteristics make it favorable for use on Navy ships and
submarines.
CFC-114 AND THE U.S. NAVY
CFC-114 has been in use on Naval ships since 1969 and has demonstrated excellent reliability. However,
design improvements have often lagged behind commercial advancements in compressor technology, advanced
heat transfer surface technology, and intelligent control system technology. It is costly and difficult to keep up
with commercial advancements when the Navy uses CFC-114 and the much larger shipping industry uses CFC-11.
The Navy, however, has made significant progress in recent years in advanced heat transfer surface technology [6].
CFC-11 was found to be unsatisfactory to the Navy because of problems unique to ship and submarine application.
For example, CFC-11 operates at sub-atmosplieric pressures and therefore is subject to air and water vapor
infiltration leading to corrosion of system components. Additionally, CFC-11 decomposes at high temperatures
causing toxicity problems on submarines as the air is recycled in high temperature air purification equipment In
contrast, CFC-114 operates at approximately atmospheric pressure and remains stable at lugh temperatures.
Some of the Navy's unique requirements include the need for small refrigerant inventory and small
components due to space constraints. Efficiency has been a low priority in the past but with shrinking defense
budgets it has become more important. Additionally, ships and submarines need to operate silently in tactical
situations and recycle air in living spaces. Coaling systems must be able to operate at as low as 10% of maximum
capacity during normal peacetime operations yet handle a dramatic increase in load when firing weapons in
combat or training situations. Fully halogenated refrigerants, such as CFC-114, generally exhibit the best
compatibility and impose the least restriction in choice of materials; a suitable replacement must display similar
material compatibility. Other requirements for a suitable replacement include meeting safety and environmental
standards for toxicity, flammability, ODP, and GWP.
6
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The Navy will likely design new cooling systems with HFC-134a. However, a need still exists for a
suitable near term replacement for existing equipment using CFC-114. Because industry's attention has been
focused on CFC-11 and CFC-12 replacements, the Navy must devote substantial resources to address the CFC-114
problem. Current potential alternatives for CFC-114 are not well developed and substantial modifications to
system equipment will likely be necessary in order to accommodate them. For example, HCFC.-124 operates at
much higher condenser pressures than CFC-114 requiring impeller and heat exchanger modifications.
FLUORINATED HYDROCARBONS
At one time, HCFC-123 and HCFC-124 were leading alternatives for CFC-11 and CFC-114, respectively.
When it became apparent that these HCFCs would also be phased out as environmentally unsuitable, the EPA
began investigating "back up" alternatives [7]. As a result, a series of propanes have emerged as candidate
replacements for CFC-114.
The EPA set selection criteria that considered thermodynamic properties, GWP, ODP, ease of
manufacture, toxicity, and flammability and then decided to pursue fluorinated ethers and fluorinated propanes
One of the replacement candidates screened, namely HFC-236ea, is the focus of this study. HFC-236ea appears to
be less toxic than CFC-114, is miscible with polyolester oils, is not flammable, has a 1.2-year atmospheric lifetime
and has a known method of production from hexafluoropropylene. Initial modeling by the EPA predicts
performance to be within 1% of CFC-114. However, prior to the present study there were no data available for the
performance and heat transfer characteristics of HFC-236ea in a typical shipboard chiller Design changes to the
Navy's existing equipment will likely be required in order to accommodate HFC-236ea or any other alternative
refrigerant. A simulation of a typical shipboard chilled water system was therefore deemed useful for future design
and optimization of Navy chillers.
SUMMARY
The United States Navy presently uses CFC-114 as the working fluid in water chillers used for electronics
and space cooling. With a mandatory phase-out of CFCs in place, it is necessary to replace CFC-114 in these
shipboard chillers with an alternative refrigerant that does not contribute to ozone depletion or global warming. Of
special importance to the Navy is finding a replacement refrigerant that is non-toxic because of the closed
environments aboard ships and submarines. In addition, energy efficiency is important because space-consuming
fuel must be carried aboard ships during deployment. Finally, reliability and material compatibility are important
for the replacement refrigerant because of the need for combat readiness and the fact that ships are commonly
deployed away from repair facilities.
7
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HFC-236ea is a promising candidate for replacing CFC-i 14 for several reasons. First, unlike other
replacements such as E-134, there is currently a commercial production route available for large quantities through
the use of hexafluoropropylene. Second, initial modeling conditions appear very favorable as a drop-in substitute,
with modeled performance being within ]% of CFC-114 and operating capacities, pressures, and temperatures
matching closely [7]. Flammabilily tests, materials compatibility tests, and oil miscibility tests appear highly
favorable. Preliminary results indicate that HFC-236ea is miscible with a commercial polyolester oil and is not
flammable. Material compatibility testing confirms HFC-236ea and a polyolester oil in the presence and absence
of water to be compatible with aluminum, steel, copper, Mylar, Nomex, Viton and Buna-N Acute inhalation test
results indicate lower acute toxicity than CFC-114, which minimize long term effects on the environment. In
addition, estimates predict that HFC-236ea has a short atmospheric lifetime.
8
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CHAPTER 2. CONCLUSIONS
The Montreal Protocol began a worldwide drive to eliminate the production of chloroflourocarbons which
are thought to be harmful to the environment As a result of the restrictive legislation that followed, there is an
immediate need to replace CFC-114 which is used extensively on United States Navy's ships and submarines.
Preliminary research conducted by the United States Environmental Protection Agency suggested that HFC-236ea
might perform suitably as a near term drop-in replacement for CFC-114. However, at the time of this study, heat
transfer data for HFC-236ea were not available.
For this reason, a computer model was developed for comparing these two refrigerants in a simulated 125-
ton centrifugal chiller system representative of those found in the U.S. fleet. The model is semi-empirical,
combining thermodynamic and heat transfer theory, as well as boiling and condensing heat transfer coefficient
data measured at the Iowa State University Heat Transfer Test Facility.
The Naval Surface Warfare Center in Annapolis, Maryland also provided data for this study. A 440-
kilowatt laboratory centrifugal air conditioning plant and HFC-236ea were used for the data collection. The
experimental data provided by the Naval Surface Warfare Center were compared with the modeled predictions
The model was tested for a range of inlet condenser water temperatures, entering and leaving chilled
water temperatures, and evaporator and condenser water flow rates. The simulation model predicts that HFC-
236ea would perform favorably as a drop-in substitute for CFC-114.
Additionally, several recommendations were provided for improved performance using HFC-236ea in
centrifugal chiller systems Design recommendations discussed in this study include manipulating the evaporator
load to achieve positive gage refrigerant pressure, ensuring the absence of non-condensable gases in the system,
using a variable speed compressor with a fixed inlet guide vane angle to the impeller, conducting further research
using azeotropic mixtures with HFC.-236ea as the major component, and installing high performance enhanced
surface tubes in both the evaporator and the condenser.
In conclusion, the simulation developed in this study provides results that are consistent with the expected
behavior of a 125-ton refrigeration system. The results provided by the Naval Surface Warfare Center sufficiently
validate the model. Finally, the results suggest that HFC-236ea would perform well in existing CFC-114
centrifugal chillers, although design modifications should be considered for optimal performance.
9
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CHAFFER 3. RECOMMENDATIONS
One way to improve the performance of the fleet 125-ton chiller and allow the use of HFC-236ea as an
alternative working, fluid is to reduce the load on the evaporator by increasing the temperatures of the chilled water
entering and leaving the evaporator by a few degrees. This would allow the refrigerant temperature in the
evaporator to come up slightly, which in turn would result in an evaporator pressure that is above atmospheric
pressure. With a positive gage pressure in the evaporator, there is less possibility of non-condensable gases and
contaminants to leak into the system where they can accumulate in the condenser and reduce performance. The
low evaporator temperatures and lugh condenser temperatures reported by the NSWC suggest the possibility of this
occurrence. This proposed solution avoids the cost of redesigning system components.
A purge device should also be installed at the highest point of the condenser to allow purging of non-
condensable gases that might accumulate there. If non-condensables is a persistent problem, the purge unit may be
malfunctioning or the system may have an air leak larger than the purge unit can handle
A variable speed compressor would eliminate the need for hot gas by-pass or the extensive use of inlet
guide vanes in the compressor to control the refrigerant flow. A variable speed chiller would allow the maximum
system performance to be realized over a broad range of operating conditions resulting in maximum energy
savings.
Another possible improvement might be to mix HFC-236ea with other non-CFC refrigerants to form an
awotropic mixture with properties that allow the saturation point in the evaporator to stay above atmospheric
pressure. The mixture could be chosen so as to maintain the positive properties ofHFC-236ea.
Additionally, better performance in the Navy's fleet air conditioning units could be realized by investing
in commercially available high performance tubes. While not reported in this study. Turbo B tubes were simulated
with CFC-114 and HFC-236ea under fleet design conditions and were predicted to perform significantly better
than 10.23 fins per centimeter tubes in both the evaporator and in the condenser.
Finally, the model predicts that HFC-236ea used as a drop-in substitute for CFC-114 without any design
modifications may result in energy savings. The model predicts that for any set of conditions, the power required
for a refrigeration cycle using HFC-236ea as a drop-in will be significantly less than the same cycle using CFC-114
as the working fluid The predicted savings in power consumption by using HFC-236ea at the design point of
operation is 8.6% which is equal to a .550 kW. If HFC-236ea is to be used only as a near term replacement, it may
be appropriate to use it without making any significant design changes to the system.
10
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CHAPTER 4. REVIEW OF RELATED LITERATURE
The vapor-cotnpression cycle is the most widely used cycle for refrigeration and air-conditioning. Past
investigators have studied the vapor-compression system from both theoretical and experimental points of view.
These past studies have been aimed at understanding the behavior of each of the components of a system.
Theoretical studies are usually carried out with the aid of modeling techniques that make use of digital computers.
As computers have become more powerful, these models have become progressively more detailed.
Many studies of the vapor-compression cycle have focused on modeling reciprocating compressors.
Threlkeld provides ail example of theory for a simple model of a compressor piston assembly that can readily be
written into computer code [8], This was accomplished in a study by Smith et al. in which several variations of a
vapor-compression cycle were modeled in an interactive computer program particularly designed for student use as
an investigative tool [9J.
Due to the complexity of fluid behavior in a centrifugal compressor, there are few reports of successful
computer modeling efforts found in the literature. Some examples are discussed in this chapter. Table 4 1
compares the models mentioned below, highlighting some of the important characteristics of each.
Braun et al. developed a mechanistic model of a centrifugal chiller operating with variable-speed capacity
control [10] The model utilizes mass, momentum, and energy balances on the compressor, evaporator, condenser,
and expansion device. Given a chilled water setpoint temperature and entering chilled and condenser water
temperatures and flow rates, the model predicts both the required compressor speed and power consumption. The
model was compared with performance data for a 5500 ton variable-speed centrifugal chiller at the Dallas/Fort
Worth airport. This model requires empirically derived constants to characterize the compressor.
A computer simulation model was developed by Jackson et al. to analyze the performance of a water-
cooled, variable-speed centrifugal chiller with hot gas bypass option for capacity control [11]. The model is based
on thermodynamic principles and empirical correlations and was calibrated using available capacity test data. The
performance of the chiller at various conditions and design modifications was predicted using the calibrated model
and results of the parametric performance study were presented. The model requires a compressor map and other
empirically derived constants.
Wong and Wang developed a model of a two-stage centrifugal chiller using a water-cooled condenser and
CFC-11 as the refrigerant [121. The heat exchangers were modeled as a shell-and-tube type. The centrifugal
chiller was driven by a hermetic motor, and capacity was controlled by the use of inlet guide vanes at the inlet of
the first and second-stage impeller. The model was structured such that the load ratio and the entering temperature
of condenser water were the two independent variables. The model depends on compressor performance maps and
other empirically obtained inputs. The results of the model were compared with actual operating results.
11
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Table 4.1: Comparison of models found in the literature
Author
Smith et al.[9]
Type of
"Mattel 'I-
Compressor
Analysis
Type of
Compressor
Type of
Condenser
Refrigerant
l2,T2, 502
theoretical
isentropic
reciprocating
no
restriction
Braun et al.[10]
semi-
empirical
polytropic
centrifugal
water
cooled
12, 22, 500,
Jackson
et al.[11]
semi-
empirical
isentropic
centrifugal
water
cooled
114
Wong arid
Wang [12]
semi-
empirical
isentropic
centrifugal
water
cooled
11
Domanski and
McLinden [131
theoretical
isentropic
reciprocating
air cooled
mixtures -
Bare [7]
theoretical
isentropic
n/a
no
restriction
propanes
and ethers
A simulation program, "CYCLE11" was developed by Domanfiki and McLinden [13], This model
simulates vapor-compression cycles in a heat pump and in a refrigerator. The model requires the input of an
average effective temperature difference representing a generalized temperature difference between the heat
transfer fluids in the heat exchangers. The model utilizes the Carnahan-Starling-DeSantis equation of state which
provides the thermodynamic properties for several refrigerants and refrigerant mixtures.
Chlorine-free fluorinated ethers and fluorinated hydrocarbons were studied by Bare as potential long-term
replacements for CFC-1! and CFC-114 [7], A model was used to predict the performance of these chlorine-free
compounds in a variety of refrigeration applications The model utilizes the Carnahan-Starling-DeSantis equation
of state and allows analysis of a simple theoretical vapor compression cycle. The model predicts that HFC-236ea
will perform within 1% of CFC-114 based on a thermodynamic analysis only. All simulations were based solely
oil thermodynamic properties and analyses, transport properties were not included in the model and heat transfer
effects were not taken into account.
12
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CHAPTER 5. VAPOR-COMPRESSION CYCLE
There are many types of refrigeration cycles that perform the function of removing heat from a region of
low temperature and discharging this heat to a region of higher temperature. Examples of these cycles include air,
steam-jet, absorption, thermoelectric, and vapor-compression refrigeration cycles. All of the above cycles have
been described and compared in detail by Gauger et al., 1995 [14], Of these cycles, the vapor-compression
refrigeration cycle is the most commonly used system in commercial and residential applications.
The vapor-compression cycle is characterized by a working fluid that is vaporized, compressed,
condensed, and expanded in a complete cycle. The basic components of this closed system are shown in Figure 5.1
and include two heat exchangers, a compressor, and an expansion device. Also shown in the figure are the work
and heat transfers, which are positive in the direction of the arrows.
a
Condenser
Expansion
Device
Compressor
IV..






Evaporator



1
tie
Figure 5 .1: Components of a vapor-compression refrigeration system
The refrigerant vapor is moved by the compressor to the condenser where it is de-superheated, condensed
and possibly subcooled by heat transfer to a circulating coolant. The liquefied refrigerant then moves through an
expansion device where the pressure is reduced and the liquid partially flashes into vapor, thereby lowering its
temperature. The two-phase mixture then flows through the evaporator, where it is fully evaporated and slightly
superheated, while absorbing heat from the fluid to be cooled by the cycle. The low-pressure refrigerant vapor
leaving the evaporator is then drawn to the compressor and the cycle is repeated.
13
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The simple vapor-compression cycle is better understood with the aid of a temperature-entropy diagram.
In Figure 5.2, a typical simple vapor-compression cycle is represented by the path 1-2-3-4-1. The compressor
receives low-pressure refrigerant vapor and compresses it adiabaticaliy and reversibly. The high-pressure,
superheated vapor enters the condenser and condenses at constant pressure to a liquid. Irreversible and adiabatic
expansion takes place in the expansion device, and the resulting low pressure refrigerant absorbs heat in the
evaporator at constant pressure to complete the cycle.
Some differences between ideal and actual refrigeration systems are briefly discussed below.
•	In an ideal cycle, the refrigerant vapor leaving the evaporator is often assumed to be saturated vapor. In an
actual cycle, refrigerant vapor leaving the evaporator is superheated a few degrees to add a safety margin in
avoiding the undesirable effects of wet compression.
•	Similarly, the refrigerant liquid leaving the condenser is often assumed to be saturated liquid. However, the
refrigerant liquid leaving the condenser is preferably subcooled.
•	Ideal cycle heat transfer processes in the evaporator and condenser are internally reversible. In an actual
cycle, however, friction causes pressure drops in the heat exchangers as well as local temperature differences.
External irreversibilities require that a finite temperature difference between heat transfer fluids and the
refrigerant exist to allow heat transfer. This is illustrated in Figure 5.2 where the warm region (heat sink) may
T
s
Figure 5.2: T-s diagram of a typical vapor-compression refrigeration cycle
14
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be a heat transfer fluid flowing through the condenser to which heat is rejected. Likewise, the cool region
(heat source) may be another heat transfer fluid circulating through the evaporator which absorbs heat from
the refrigerant
•	The ideal compressor operates reversibly and adiabatically, whereas the real compressor experiences friction,
heating, and irreversibility.
•	No state changes in the working fluid occur except in the components in an ideal cycle. In reality, pressure
drops occur in the long suction and discharge line piping resulting in increased compression work.
•	In an ideal cycle, components including the compressor and the suction and discharge lines are assumed to be
isentropic (reversible adiabatic). However, in a real cycle, heat transfer occurs between system components
and their surroundings.
•	In an idealized model and in this study, changes in kinetic or potential energies throughout the system are
assumed to be negligible.
•	Finally, ideal cycle components operate at steady state while actual systems experience transient effects.
The following is a brief discussion of the major components of a vapor-compression system including the
condenser, evaporator compressor, and expansion device. The discussion includes an explanation of the
assumptions made in modeling these components in this study.
CONDENSER
The condenser receives superheated vapor from the compressor, removes the superheat, and then liquefies
the refrigerant. Different types of condensers include air-cooled, water-cooled, and evaporative condensers A
water cooled, shell-and-tube condenser is modeled in this study. When adequate low-cost condensing water is
available, water-cooled condensers are often desirable because lower condensing pressure and better control of the
discharge pressure is possible. Water, especially when obtained from underground sources or a big heat sink, such
as the ocean, is usually much colder than daytime ambient air temperatures Because of the excellent heat transfer
characteristics of water, water-cooled condensers are usually quite compact. A shell-and-tube condenser acts as
both a condenser and a liquid receiver. It is constructed of a vessel having a refrigerant inlet and outlet An
example of a typical shell-and-tube condenser is shown in Figure 5.3.
The condenser modeled in this study is assumed to be internally reversible. The specifications provided
by the NSWC include the following: The shell consists of 246 copper finned tubes at 10.23 fins per centimeter
(26 fins per inch). The water on the tube-side makes two passes through the shell. The shell is 55.88 centimeters
outside diameter (OD) and 237.49 centimeters in length. The effective tube length is 220.31 centimeters and the
average tube outside surface area-to-length ratio is 0.64. Based on this information, the calculated outside surface
area is 105.81 square meters
15
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Hot gas inlet
Baffle plate
	
Water out
Water in
Liquid line
Figure 5 .3: Diagram of a typical shell-and-tube condenser
EVAPORATOR
The evaporator in a refrigeration system is a heat exchanger that removes heat from the space or heat
transfer fluid being cooled. A flooded shell-and-tube evaporator is modeled in this study. Flooded systems operate
with a definite liquid refrigerant level in the evaporator. This liquid refrigerant level is maintained in the
evaporator through the action of a refrigerant flow control device. There are several advantages of the flooded
system over other systems. A few of these advantages are: higher efficiency, lower operating costs, less cycling,
higher rate of heat transfer, and closer control of temperature. More liquid on the low-pressure side of the system,
as in the flooded system, provides a greater area of wetted surface and allows a higher rate of heat transfer through
the evaporator walls and tubing. An example of a typical shell-and-tube flooded evaporator is shown in Figure 5 4
The specifications provided by the NSWC for the evaporator of interest in this study are similar to those
for the condenser mentioned above. The evaporator is assumed to be internally reversible. The shell holds 246
copper, 10.23 fins per centimeter tubes. The wafer flowing through the tubes makes two passes through the shell.
The shell's outside diameter (OD) is 81.28 centimeters and its length is 237.49 centimeters. The effective tube
length is 220.31 centimeters and the average tube outside surface area-to-length ratio is 0.64. Based on this
information, the calculated outside surface area is 105.81 square meters.
16
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Liquid level
Eliminator
vapor outlet
Dropout
Tube sheet
Distributor
Refrigerant
inlet
CD
Chilled water
outlet
CD
Chilled water
inlet
Figure 5.4. Diagram of a typical shell-and-tubc flooded evaporator
COMPRESSOR
The compressor draws vapor from the suction line or accumulator, compresses it to a higher temperature
and pressure, and then discharges the superheated vapor into the condenser. Types of compressors include
reciprocating, rotary, helical rotary (screw), and centrifugal. In reciprocating and rotary compressors, the
refrigerant molecules are squeezed together inside the cylinder by the positive action of the piston or rotor.
Compression is produced and maintained by the action of the suction and discharge valves. In contrast, centrifugal
compressors are characterized by a continuous exchange of momentum between an impeller and a steadily flowing
fluid Pressure is produced when gaseous refrigerant, whirled at a high rate of speed, is tluown outward by
centrifugal force and caught in a channel. The centrifugal compressor is the dominant type of compressor used in
large installations and is the type modeled in this study.
Centrifugal compressors are in the family of turbomachines, which also include fans, propellers, and
turbines Because their flows are continuous, they have large volumetric capacities. Multiple stages can be
installed to increase the pressure lift of the compressor. Centrifugal compressors are efficient and well suited for
large capacity refrigerating plants ranging from 175 to 10,500 kilowatts [15]. They are efficient at a wide range of
operating temperatures. Because they are not positive-displacement type compressors, they are flexible under
varying load conditions and operate at good efficiencies even when the demand is less than 40% of their designed
capacity [16].
17
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Although centrifugal compressors require high rotative speeds, there is minimal wear and vibration due to
the lack of contact between moving parts. Lubrication is not needed at any place on the centrifugal compressor
except at the end bearings of the shaft. Since these end bearings are the only internal friction surfaces, the
refrigerant vapor compressed by a centrifugal compressor is free from oil, giving it the advantage of preventing an
accumulation of oil on the heat transfer surfaces of the condenser and evaporator.
The compressor modeled in this study is assumed to follow a reversible polytropic process. A constant
polytropic exponent, n - 1.04, is assumed based on an average value calculated from a representative sample of
performance data provided by the NSWC. The compressor specifications provided by the NSWC include the
following: the compressor is an open, single-stage, centrifugal compressor-motor driven unit in a refrigeration
system having 125 tons of cooling. The compressor is direct-driven through a torque meter station and operates at
an impeller speed of 11,918 revolutions per minute through an internal compressor gear arrangement.
EXPANSION DEVICE
The expansion device regulates the flow of refrigerant from the high-pressure to the low-pressure side of
the system. Some common types of flow control devices include orifices, capillary tubes, high pressure float
valves, and thermostatic expansion valves. Capillary tubes are passive devices, common for small applications
such as domestic refrigerators. A thermostatic expansion valve controls the degrees of superheat at the evaporator
outlet. An adjustable orifice is modeled in this study as a throttling process.
An orifice is a refrigerant flow control device used to control the refrigerant level in the flooded
evaporator. Orifices and capillary tubes perform basically the same function, and they are practical only for
systems which operate at nearly constant capacity. They are sized to pass refrigerant liquid at a slightly greater
rate than desired for the pressure difference available. This results in exhausting the liquid supply in the
condenser. The disadvantage of this type of control is that it allows some gas to leave the condenser and carries
additional enthalpy to the evaporator. This loss is not large in a reasonably constant capacity system. The low-
cost, simplicity, and dependability of this type of liquid feed control more than compensates for its slight
inefficiency in centrifugal chilled water systems. For the purpose of this study, it is assumed that the oriface is
ideal and that the liquid leaving the condenser is slightly subcooled.
18
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CHAPTER 6. MODEL DESCRIPTION
A computer program has been developed that simulates the performance of a 440-kilowatt capacity,
single-stage, centrifugal, chilled water air-conditioning plant. The design conditions shown in Table 6 1 are based
on the design of a typical air-conditioning plant in use on Navy ships and submarines.
Table 6.1: Simulated design conditions
Component
Design Condition
Value
Evaporator
chilled water flow rate
28.4 l/s
Evaporator
entering chilled water
temperature
10.7 °C
Evaporator
leaving chilled water
temperature
7.0 °C
Condenser
water flow rate
31.5 l/s
Condenser
entering water
temperature
31.4 °C
Given the entering and leaving temperatures of the chilled water, the entering temperature of the
condenser water, and the flow rates of the chilled water and condenser water, the model predicts the required
compressor power and the saturation temperatures in the heat exchangers. With knowledge of fluid properties and
tube geometries, the performance of the system with different refrigerants and enhanced surface tubes can be
compared under similar operating conditions. For the purpose of this study, the model is used to compare
refrigerants CFG-114 and HFC-236ea using 10.23 fins per centimeter tubes in the condenser and evaporator The
results are presented in the next chapter.
The model allows imposed evaporator superheat and imposed condenser subcooling. If wet compression
is encountered during the iteration procedure, the model adjusts the degrees of superheat just enough to stay in the
dry region. The compressor is modeled using a polytropic analysis [17]. The polytropic exponent is estimated
from data provided by the Naval Surface Warfare Center and is assumed constant at n = 1.04. The externally
adiabatic heat exchangers are assumed to be internally reversible. Heat transfer coefficients are provided from heat
flux data taken on a single tube testing facility at Iowa State University. The throttling process is assumed to be
adiabatic and irreversible. An iteration procedure is used to solve for the evaporator and condenser saturation
temperatures
19
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PROPERTIES
Properties for this simulation are estimated using subroutines from a computer program, "REFPROP"
version 4.01, developed by the National Institute of Standards and Technology [18] The utility of REFPROP
includes the ability to estimate thermodynamic and transport properties for refrigerant mixtures of up to five
components using the Carnahan-Starling-DeSantis equation of state. For pure refrigerants, such as CFC-114 and
HFC-236ea, the REFPROP subroutines calculate properties using an extended corresponding states model In this
model, the properties of a range of related fluids are scaled to a well characterized reference fluid, HFC-134a. The
transport properties of thermal conductivity and viscosity, which are important in calculating heat transfer
coefficients in the heat exchangers, are also calculated using an extended corresponding states model [IS, 19],
This study is limited to a comparison of two pure refrigerants, namely CFC-114 and HFC-236ea.
However, by using REFPROP subroutines, other pure refrigerants or refrigerant mixtures supported by REFPROP
may be simulated as the working fluid in the refrigeration system of this study within the constraints of the laws of
thermodynamics. Thus, it is a useful tool in evaluating alternative non-CFC refrigerants in existing systems and
would also be a useful tool in future simulations of innovative vapor-compression cycles
PROGRAM DESCRIPTION
A complete listing of the Fortran code for the main program is included as Appendi x A. A description of
the main program structure is described in the paragraphs that follow. Throughout the description, both the
numerical procedure and the modeling theory are discussed in detail. Figure 6.1 is a flow diagram of the computer
program. Common blocks and dimension statements are set up for use with REFPROP subroutines and Newton-
Raphson subroutines A series of data statements and input prompts are used to identify the design conditions,
initial temperature estimates, and tube geometries REFPROP subroutines are then initialized by identifying the
number of components, mixture composition, component names, and the choice of computational model
Additionally, the reference values for enthalpy and entropy are selected. Subroutine BCONST is then called to set
up equation-of-state parameters from stored property data. All units for REFPROP subroutines are specified using
the International System of Units (SI). After the program is initialized, simulation conditions arc selected and
varied over an appropriate range of operating conditions. Default values for the model are equivalent to fleet
design conditions and are listed in Table 6.1.
20
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begin
no
solution?
yes
next
value
data file
end
select
refrigerant
output
results
solve thermo
cycle
select range
calculate
hx coefficients
initialize property
routines
select parameter
to vary
estimate
hx temperatures
Figure 6.1: Program flow diagram
21
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THERMODYNAMIC ANALYSIS
Each of the four basic components of a vapor-compression system—namely the compressor, the condenser,
the expansion device, and the evaporator—has its own peculiar behavior. At the same time, each component is
influenced by conditions imposed by the others For example, a change in the condenser water temperature may
change the rate of refrigerant flow, which in turn may cause the heat exchanger temperatures and pressures to
change as well as change the power required to the compressor. This study models the individual components of
the vapor-compression cycle and also observes how they interact with each other as a system.
Pressure (P) and enthalpy (h) are two properties that may conveniently represent a vapor-compression
system. A P-h diagram for refngerants CFC-114 and HFC-236ea is shown in Figure 6.2. Simple vapor-
compression cycles for CFC-114 and HFC-236ea at fleet design conditions are also shown in Figure 6.2. Four
state points for each cycle are identified respectively as: (1) evaporator outlet and also compressor inlet, (2)
compressor outlet and also condenser inlet, (3) condenser outlet and also expansion device inlet, and (4) expansion
device outlet and also evaporator inlet Often, when modeling a simple vapor-compression cycle, superheat at the
evaporator outlet (state 1) and subcooling at the condenser outlet (state 3) are either imposed or assumed to be zero.
Superheat at the compressor inlet is normally desired to avoid the occurrence of "wet compression" which
degrades system performance and may cause damage to the compressor impeller over time Subcooling at the
condenser outlet is beneficial to the performance of the system because it allows a greater enthalpy difference
across the evaporator resulting in greater cooling capacity. Constant pressure is often assumed in both heat
exchangers and in the suction and discharge lines to the compressor, although in reality the irreversible nature of
the processes in these components will result in slight pressure differences.
With initial estimates of saturation temperatures for the evaporator and condenser, all remaining
thermodynamic and transport properties of interest can be calculated for the cycle modeled in this study. Much of
the theory for the thermodynamic analysis that follows is discussed in more detail in Moran and Shapiro [20],
State point (1) is defined as saturated vapor at the estimated refrigerant temperature in the evaporator.
The state of the pure refrigerant is thereby fixed, and the remaining thermodynamic properties of interest including
pressure, enthalpy, and specific volume are calculated using property subroutines.
Pl=fJ(Te)	(6.1a)
(6.1b)
v'l =fi(Te)	(6.1c)
Similarly, the outlet of the condenser is defined as state point (3) and is fixed by the estimated temperature
of the condenser at the saturated liquid point. The pressure and enthalpy are then determined with equation of
state calculations.
22
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r>3 = f4 (rc:)
*3 ~/5(7'c)
(6.2a)
(6.2b)
10000
-------
The pressure at state point (4) corresponds to the estimated saturation temperature of refrigerant in the
evaporator, and the slate is fixed in the two-phase region. The basic thermodynamic properties of each stale point
of the refrigeration system are known based on estimated saturated temperatures of the evaporator and condenser,
and on a constant polytropic exponent obtained front performance data. Energy balances are used along with
minimum input design conditions to iteratively solve for the evaporator and condenser saturation temperatures.
Energy Balance
The evaporator load is determined by the first law relationship:
Qe ~ (mch\v X ^pc)tw X lee ~ ^ el)	(^.6)
where the flow rate of the chilled water, mcjv^ along with the entering and leaving temperatures of the chilled
water, Tee and 7e/, are known. The specific heat, Cp f is calculated as a function of the average chilled water
temperature, Tc}n,h
—	I pi* ^ t ' /
Tchw = 2	.	(6 7)
Fresh water and sea water properties are calculated using property subroutines so that temperature effects
are taken into account in calculating specific heats
^'Pchw ~ f^chw)	(6-8)
Next, the mass flow rate of the refrigerant is calculated as:
Qe
_	(6.9")
' hl~h4
and the heat removal rate in the condenser as:
Qc = mr ( /?2 — ^3 )	(6.10)
Finally, making use of knowledge of the flow rate of sea water in the condenser and the temperature of the
entering sea water:
Tel Tce +	)(C~ )	(611)
1 mSW A	)
-	Tce 4- Tci
7™ = 2	^612)
~ /( Tyw )	(6.13)
24
-------
Since estimates of Tc and Te have been used up to this point, they represent two unknowns in the set of equations
for which two additional equations are needed to make the set complete. These equations are:
Qc — UAoediTime	(6.14)
Qc ~~ L'^o.c ^lm,c	(6 15)
where the log mean temperature for the evaporator is defined as:
A7W = In (Tce - T*) - \n(Te! - Te)	(61<3)
and the log mean temperature for the condenser is:
A7W = In (Tc - T^)~ 1 n(Tc - Tcj)	(6 ll)
and all the temperatures are either known or estimated.
HEAT TRANSFER ANALYSIS
The overall heat transfer coefficient for the heat exchanger, U, is multiplied by the total outside surface
area, Afj, of the heat exchanger The overall heat transfer coefficient is a function of temperature and other fluid
properties, and it can be predicted using a variety of published correlations.
Evaporator
The evaporator modeled is a shell-and-tube type with 246 tubes. The tube-side water makes two passes
through the heat exchanger. The actual cooling surface (outside tube) area varies with the tube type, but it is
known to be 105.81 square meters for 10.23 fins per centimeter tubes. The shell is 81.28 centimeters OD and
237.49 centimeters in length. The design conditions for the evaporator, as stated previously, are chilled water
flow rate equal to 28.4 liters per second and inlet and outlet chilled water temperatures of 10.7 °C and 7 °C,
respectively. The heat transfer analysis begins by examining the heat exchanger geometry. The total outside tube
surface area, Aa e, of the evaporator is equal to the number of tubes times the average outside surface area per
length of the tube times the effective tube length-per-pass, in the heat exchanger. Note that the average outside
surface area of the tube, AQ e , is expressed as surface area per unit length (m2/m) which includes the actual
surface area of the fins. The value used was provided by the NS WC.
A
-------
The inside surface area of the evaporator tube (smooth) is calculated as the total number of tubes times the
mside perimeter times the effective tube length-per-pass through the heat exchanger:
Aj e — A'e n(die )U'e£fg )	(619)
The total flow area of the evaporator tubes is equal to half the total number of tubes (two pass) times the
cross-sectional area of the tube:
Ne n(d; p)~
=	(62°)
The average water velocity in an evaporator tube is then calculated as:
7, . 	Iklm		«,n
Cpc,„»!/,„)	('l}
where the average water density is based on the average chilled water temperature in the evaporator. The average
Reynolds number follows as:
Rechw = <^hwXyehwXllU) .	(6 22)
t'chw
and average Prandtl number as:
fW =	(6.23)
Kchw
The inside heat transfer coefficient and the tube wall temperature are then calculated by iteration. The
average wall temperature is initially assumed to be equal to the saturated refrigerant temperature in the evaporator:
= Te	(6-24)
The average viscosity of water is then calculated at the estimated temperature:
T1m\e ~~ fW\v,c )	(6.25)
The inside or water-side heat transfer coefficient for the evaporator is calculated using the Dittus-Boelter
correlation [21]:
k
•r"i
flw,e
Ke : 0023 i^chw )°-8 (P'-chw )°3 ( )°-M ( 7?* )	(6 26.)
All water properties are calculated at the average chilled water temperature unless otherwise subscripted.
The wail temperature of the tube is then recalculated as:
Tw,e = W-(/7jjiAje)	(6 27)
26
-------
The outside boiling coefficient of the evaporator, ha e, is calculated from measured heat flux data
provided from a separate study conducted at the Iowa State University Heat Transfer Test Facility [22] Boiling
coefficients for CFC-114 and HFC-236ea using a single tube 10.23 fins per centimeter test rig were calculated as a
function of heat flux and constant saturation temperature. For HFC-236ea and 10.23 fins per centimeter the
correlation is given as:
ho e = 2.22792 l 0.1742529 (qe") - 1.766886E"3 (qe")2	(6.28)
and for CFC-114 and 10.23 fins per centimeter the correlation is given as:
K>,e = 0.8431786 + 0.1359888 (qe") - 8.738483E"4 (.qe")2	(6.29)'
Finally, the overall heat transfer coefficient neglecting thermal resistance of tube wall is calculated as:
UAo,e = [ ( hieUlfi) ' (h0^(Ao e)1 ¦	(6'30)
Condenser
The condenser is modeled as a shell-and-tube falling film condenser. The shell is 55 88 centimeters OD
and 237.49 centimeters in length. The heat exchanger is designed for sea water at a flow rate of 31.5 liters per
second with an entering temperature of 31 4CC, making two passes through the shell. There is a total of 246 tubes
in the shell and a heat exchanger outside surface area of 105,81 square meters for 10.23 fins per centimeter tubes
The area varies with tube type.
The total outside tube surface area, A0 c , of the condenser is equal to the number of tubes, A'c, times the
average outside surface area per length of the tube, A0 c, times the effective tube length per pass,
heat exchanger:
-V - 'Vc^KV/.c)	("')
The inside surface area of the condenser tubes is calculated as the total number of tubes times the inside
perimeter times the effective tube length per pass through the heat exchanger:
Auc = Aicx(dKC)(.LejLc)	(6.32)
The total flow area of the condenser tubes is equal to half the total number of tubes (two-pass) times the
cross-sectional area of a tube:
Nc n(d; c)2
Afc=^[—^~]	(6.33)
27
-------
The average water velocity in the condenser tubes is then calculated as:
-	f,lnv
V =	® —	(6 34)
w (PsvMfr)	¦
where the average water density is based on the average sea water temperature in the condenser. The average
Reynolds number follows as:
Z' _ (PsvM Kmdidj.e)	,fi
Rcsw ~	7.	
-------
and for CFC- l 14 and 10.23 fins per centimeter the correlation is given as:
~ho c = 3.620498 + 0.1494268 {qc") - 2.087891E"3 {qc")2
(6.42)
The overall heat transfer coefficient is calculated as:
(6.43)
NUMERICAL PROCEDURE
The equations above may be solved simultaneously with the use of a computer. The method chosen for
this simulation is the Newton-Raphson procedure for multiple equations and unknowns based on a Taylor-series
expansion. This method is explained in more detail by Stoecker [23], The unknown values in the above set of
equations are essentially Tc and Te, which are present in both the log mean temperature and the UA calculations.
The basic steps to the Newton-Raphson procedure are as follows:
1)	Solve as many of the equations outside of the iteration scheme as possible.
2)	Identify the remaining equations to solve using the Newton-Raphson method
3)	Rewrite the equations so that all of the unknown terms are on one side of the equality sign.
4)	Assume initial values for the variables.
5)	Calculate values of fl through ffr at the temporary values (this becomes the B matrix). The functions are
stored in a separate function routine allowing the main program to remain flexible for use in solving future
problems.
6)	Compute partial derivatives of all functions with respect to all variables (this becomes the A matrix). This
procedure is accomplished by repeated calls to a function routine that numerically calculates the derivatives
of the input functions with respect to the input variables.
7)	Using L1NPACK routines, the set of equations AX - B can now be solved where "A" is the matrix of partial
derivatives and "B" is the matrix with values of functions using temporary values of unknown variables. "X"
is equal to the difference between the temporary values and the correct values of the variables.
8)	Update the values of the variables.
9)	Test for convergence (within 0.001 for all variables).
10)	When the routine has met the established convergence criteria for all of the variables, return to the main
program where final calculations of interest may be performed and the results printed to a file.
29
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CHAPTER 7. RESULTS AND DISCUSSION
To evaluate HFC-236ea as a potential drop-in replacement for CFC-114 in existing shipboard chillers, it
is useful to examine predicted performance of both refrigerants under the same operating conditions. It is also both
interesting and necessary to compare the modeled performance of both refrigerants with actual performance data.
Finally, comparisons of individual component performance may provide additional insight into the suitability of an
alternative refrigerant.
The model developed in this study allows comparisons to be made using different refrigerants as well as
several different fin tube types. Refrigerant property routines developed by the National Institute of Standards and
Technology support both pure refrigerants HFC-236ea and CFC-I14. Single tube heat transfer data for plain,
10.23 fins per centimeter, 15.75 fins per centimeter, and Turbo B tubes were provided by the Iowa State University
Heat Transfer Test Facility [22], The data include boiling and condensation heat transfer coefficients as functions
of heat flux for a given saturation temperature for both CFC-114 and HFC-236ea. Both property data and fin tube
data were incorporated into the model. Thus, by specifying the refrigerant and tube geometry at an initial prompt
in the computer routine, the model could be exercised repeatedly to simulate different refrigerants operating under
the same conditions.
COMPARISON OF MODEL AND EXPERIMENTAL RESULTS
The Naval Surface Warfare Center (NSWC), in cooperation with the United States Environmental
Protection Agency (EPA), has tested CFC-114 and HFC-236ea in a 440-kilowatt laboratory centrifugal chiller
representative of those used in the United States Navy's fleet of ships and submarines. The laboratory chiller is
fully-instrumented, and sample data are included in Appendix B.
Figure 7.1 is a comparison of modeled and measured compressor power. The measured compressor power
provided by the NSWC was calculated from measurements of torque and speed of the compressor shaft. Therefore,
the measured value of compressor power is the shaft power. The modeled value of compressor power is the rate of
work performed directly on the fluid and does not include the mechanical or heat losses as the power is transferred
from the compressor shaft to the impeller and ultimately to the working fluid. A linear relationship was found to
exist between the shaft power and the power transferred to the fluid for the data provided by the NSWC. This
correlation, Equation (7.1), was applied to the results of the model as an assumed mechanical efficiency factor
where X is the energy transfer rate to the refrigerant, and )' is the measured compressor shaft power.
Y= 391.46+ 1.0637 *X	(7.1)
30
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Even with an efficiency factor applied, the model consistently underpredicts the amount of compressor
power required to meet the specified load. This could be related lo the use of inlet guide vanes to the compressor
which are not modeled in this study.
i AA
A A	A

A HFC-236ca
uCFC-114
20	40	60	80	100 120 140 160 180 200
Modeled Compressor Power, kW
Figure 7.1: Comparison of modeled and measured compressor power
Figure 7 2 is a comparison of the system coefficient of performance, COP, calculated using NSWC
measurements and predicted using the model developed in this study. The model overpredicts the coefficient of
performance for both CFC-114 and for HFC-236ea. The trend is consistent and is what one would expect when
comparing modeled with measured results. Since models often make use of simplifying assumptions, the results
tend to be idealized One would expect to see the test results to be less favorable than modeled results.
VI
-------
^	jA	A A
A HPC-236ea
~ CFC-114
' ¦—» «
3	4	5	6	7
Modeled Coefficient of Performance
10
Figure 7.2: Comparison of modeled and measured coefficient of performance
The measured performance data appear to be fairly constant with COP values of approximately four It
appears from the data that as the temperature of water entering the condenser decreases, the difference between the
predicted and measured values of the coefficient of performance increases. The trend can be seen with HFC-236ea
as shown in Table 7.1. Measured data for an entering condenser water temperature of approximately 31.4 UC are
closer to the predicted values. However, as the temperature of the water entering the condenser decreases, the
measured coefficient of performance values increase at a slower rate than predicted values.
32
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As the temperature of the cooling water entering the condenser decreases, the heat transfer in the
condenser is enhanced due to the increased temperature difference, and the cooling capacity increases. It should
follow that overall system performance improves. However, inlet guide vanes in the compressor are used to control
the flow of refrigerant and balance the system without reducing the speed of the compressor shaft. This causes the
compressor to be less efficient and counters the effects of improved condenser performance on the overall
coefficient of performance. The model developed in this study does not account for the effects of inlet guide vanes
Thus, when inlet guide vanes are in use—for example, when the entering condenser water temperature is below the
design point—one would expect to see greater differences between measured and modeled results as shown in
Figure 7.2.
Table 7.1: Effect of entering condenser water temperature on measured and modeled COP for Hf C-236ea
Entering
iliQE
COP
modeled/
Condenser
modeled
measured
measured
Water Temp, °C



88.9
4.05
3.89
1.04
86.1
4.22
3.69 i
3.96
1.14
80.2
4.91
1.24
79.9
5.23
3.83
1.35
67.1
6.55
4.28
1.53
60.4
7.66
4.57
1.68
Figure 7.3 is a comparison of modeled and measured refrigerant temperatures in the condenser. The
measured temperature is the saturation temperature corresponding to the measured liquid pressure of the
refrigerant. The model predictions compare well with the CFC-114 data; however, the model underpredicts the
condenser temperature for most HFC-236ea data. This difference could be caused, in part, by poor heat transfer in
the condenser. If this were the case, then the temperature in the condenser would have to increase in order to
overcome whatever resistance is present. More compressor power would be required to provide this additional
temperature lift resulting in lower system performance. Because the refrigerant temperature in the condenser is
closely tied to the condensing heat transfer coefficient by use of the log mean temperature difference equation for
heat transfer in the condenser, one would also expect to see an offset in a comparison of measured and modeled
condensing coefficients. As the condensing temperature increases while entering and leaving water temperatures
remain constant, the log mean temperature difference increases. This would result in a modeled decrease in the
condensing coefficient.
33
-------
60
50
Q.
A HFC-236ea
~ CFC-114
10
0
10
20
30
40
50
60
Modeled Condenser Temperature, °C
Figure 7.3: Comparison of modeled and measured condenser temperature
An example of when conditions may exist in the condenser that hamper heat transfer is when non-
condensable gases, left unpurged, accumulate in the upper vapor space of the condenser. This is a plausible
explanation for the difference in the condenser saturation temperatures observed in Figure 7.3. As previously
mentioned, by fixing the inlet and outlet chilled water conditions as well as the chilled water flow rate, the
saturation temperature of the refrigerant in the evaporator is determined by the overall heat transfer equation
(Equation 6.14). For HFC-236ea, both the measured and modeled evaporator temperatures are near 2 UC. The
corresponding saturation pressures for these saturation temperatures are less than the atmospheric pressure. This
could cause non-condensable gases to leak into the evaporator due to the negative gage pressure. These gases
would migrate and collect in the condenser and could significantly degrade the performance of the condenser and
the entire system. If air, in fact, was present in the condenser, it would drive the outside heat transfer coefficient
down resulting in a high condenser saturation temperature.
To avoid or minimize this problem, the system should be thoroughly leak-checked and a purge installed in
the condenser. An alternate solution is to avoid negative gage pressure in the evaporator by manipulating the
chilled water mass flow rate and the chilled water temperature difference so that the saturated temperature of the
34
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refrigerant in the evaporator is raised to a minimum temperature corresponding to a saturation pressure of at least
normal atmospheric pressure. This could be accomplished with minimal effect on the evaporator capacity but
would depend on the flexibility of the shipboard heat exchangers utilizing the chilled water.
Figure 7.4 is a comparison of modeled and measured condenser capacity. The model slightly
underpredicts condenser capacity for both refrigerants The trends are consistent with Figure 7.5.
800
700
g 600
£- 500
-o 300
200
100
0
100
200
300
400
600
700
500
800
Modeled Condenser Capacity, kW
Figure 7.4: Comparison of modeled and measured condenser capacity
Figure 7.5 is a comparison of modeled and measured cooling water temperatures leaving the condenser.
Modeled values are within 0 5°C of measured values. This is consistent with Figure 7.4 which shows the same
trend for condenser capacity. This is expected, since the rate of heat transfer and the temperature of the water
leaving the condenser are the two variables in the water-side heat transfer equation for the condenser.
Figure 7.6 is a comparison of modeled and measured evaporator saturation temperatures. The figure
shows that modeled and measured boiling coefficients compare well with some variance One would therefore
expect to see a variance of the measured and modeled boiling coefficients since these variables must balance in the
log mean temperature difference equation for the heat transfer in the evaporator.
35
-------
50
40
"O
o>
A HFC-236ea
OCFC 114
0	5	10	15	20	25	30	35	40	45	50
Modeled Leaving Condenser Water Temperature, X
Figure 7.5; Comparison of modeled and measured leaving condenser water temperature
10
o
CL
£
-2
A HFC-230ea
~ CFC-114
7
0	2	4	6
Modeled Evaporator Temperature, "C
10
Figure 7.6: Comparison of modeled and measured refrigerant saturation temperatures in the evaporator
36
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Figure 7.7 is a comparison of modeled and measured refrigerant flow rate. The model consistently
predicts the flow rate for both refrigerants within ±5 percent. This suggests thai the enthalpy differences also
compare favorably since the rate of heat transfer in the evaporator is constant and is equal to the refrigerant mass
flow rate times the enthalpy difference across the evaporator.
6
cn
4
u_
"O 2
AHFC-236sa
a CFC-114
0
0
Modeled Refrigerant Flow Rate, kg/sec
Figure 7.7: Comparison of modeled and measured refrigerant flow rate
COMPARISON OF CFC-114 AND HFC-236ea PERFORMANCE
The previous figures and discussion have served to validate the model developed in this study, and it is
appropriate to further exercise the model to predict the performance of both refrigerants through a range of
operating conditions This is done by using the fleet design point as the default and varying one parameter at a
tune over an appropriate range to see the effects on the system. The results yield additional insight as to the
possible suitability of HFC-236ea as a drop-m substitute for CFC-114.
37
-------
Entering Condenser Water Temperature
As the Navy operates its fleet around the world, ships encounter a wide range of condenser water
temperatures because sea water is used directly in the condenser to remove heat from the working fluid. Chillers
for Navy ships are designed for a condenser water temperature of 31.4 °C; however, temperatures encountered may
range from -1.3 °C to 35.3 °C depending on where the ship is operating. Since heat transfer in the condenser is
driven by the temperature difference between the sea water (coolant) and refrigerant, a condenser water
temperature that is too high would lower the performance of the condenser and subsequently the entire
refrigeration cycle. Thus, the entering condenser water temperature is important to the performance of the overall
system.
In this simulation, the evaporator load is kept constant, simulating the design conditions of chilled water
entering and leaving the evaporator at 10.7 °C and 7 CC, respectively, and a constant chilled water flow rate of
28.4 liters per second. Additionally, the condenser water flow rate is held constant at the design condition of 31.5
liters per second. As the temperature of the water entering the condenser is varied, a solution is obtained and may
be expressed in terms of performance parameters such as the compressor power required or the coefficient of
performance.
For example, Figure 7.8 illustrates that the predicted power required to drive the compressor more than
doubles for both refrigerants as the water temperature entering the condenser increases from 16 °C to 38 °C. The
increasing power input trend is expected since better heat transfer and increasing heat rejection in the condenser
occurs as the temperature of the cooling water entering the condenser decreases. The efficiency of the refrigeration
cycle should thereby improve resulting in less power input required to the compressor. Figure 7.8 shows a trend
for botli refrigerants of increased power required with increased temperature of the entering cooling water to the
condenser. Additionally, the model predicts that HFC-236ea used as a drop-in substitute for CFC-114 may result
in energy savings. Figure 7.8 shows that at the design point of operation the predicted power required to drive the
compressor using HFC-236ea is 91.4 percent of the power required using CFC-114. The model predicts that for
any cooling water temperature the power required for a refrigeration cycle using HFC-236ea as a drop-in will be
significantly less than the same cycle using CFC-114 as the working fluid. The predicted savings in power
consumption by using HFC-236ea at the design point of operation is 10 kW.
Data from the NSWC are also shown on Figure 7.8. The data for CFC-114 show nearly constant
compressor power input over the range of entering condenser water temperanires. The data for HFC-236ea show
significant scatter. Both the CFC-114 and HFC-236ea measured values of required compressor power are above
the predicted values for the range of entering condenser water temperatures. When a centrifugal chiller is using
more energy than it should as suggested by Figure 7.8, a common culprit is excess air in the condenser. This
condition increases the pressure in the condenser and forces the compressor to work harder to maintain the
required cooling.
38
-------
160
140
120
1
100
tA
80
a.
E 60
o
o
A HFC-236ea data
n CFC-114 data
	CFC-114 26fpi
- - HFC~236ea2Sfpi
40 -
20
15
20
25
30
35
40
Entering Condenser Water Temperature, °C
Figure 7.8. Dependence of compressor power requirement on entering condenser water temperature
The coefficient of performance, COP, is a ratio of the cooling capacity of the evaporator over the net
power input to the compressor and is a standard measure of the performance of a refrigeration cycle. Figure 7.9
shows the refrigerating coefficient of performance as a function of the temperature of the cooling water entering
the condenser. As expected, the coefficient of performance is shown to decrease as the inlet condenser water
temperature increases. Additionally, at the design point of 31.4 °C, the predicted coefficient of performance for
HFC-236ea is 4.25 compared to 3.91 for CFC-114. The model predicts better performance using HFC-236ea over
the range of condenser water temperatures simulated.
The measured values of coefficient of performance for both CFC-114 and HFC-236ea are less than the
predicted values. As the entering condenser water temperature increases, the measured and predicted values of the
coefficient of performance move toward better agreement.
39
-------
9
8
7
c
c 6
£
k.
0
1	5
CL
O
c 4
o
'o
£
o 3
O
2
1
0
15	20	25	30	35	40
Entering Condenser Water Temperature, °C
Figure 7.9: Dependence of refrigerating performance on entering condenser water temperature
Another parameter that gives insight into the performance of a refrigeration cycle is the refrigerating
efficiency which is defined as the ratio of the coefficient of performance of the modeled refrigeration cycle to the
coefficient of performance of a reversed Carnot cycle operating between the same source and sink temperatures, in
a sense, this parameter gives a clearer picture of the cycle's true performance because it is referenced to the cycle's
best possible performance as limited by the Second Law of Thermodynamics.
The curves in Figure 7.10 show an increase in refrigeration efficiency that approach an asymptotic limit
as the temperature of the cooling water entering the condenser increases. The performance of the system at
temperatures lower than the design point is less than the possible performance which could be achieved under
those conditions. This is reasonable considering that the system being modeled was originally designed for
optimum performance at an entering condenser water temperature of 31.4 °C. The possibility that lower
temperatures result in lower refrigerating efficiencies is not of great concern for Navy applications since the
cooling fluid-in this case ocean sea water—is essentially free. The system may be designed for optimal
performance about an average cooling water temperature of 31.4 °C and any temperature encountered which is less
than that will provide extra cooling potential at no extra cost. For this reason, it doesn't have to perform optimally
40
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Fn any case, the efficiency doesn't decrease by much and, more importantly, it remains stable at higher
temperatures. Additionally, for all temperatures modeled, HFC-236ea outperforms CFC-114. At the design point
of 31.4 °C, the refrigerating efficiency of HFC-236ea is 0.332 and for CFC-114 is 0.294.
Trends in Figures 7.9 through 7.10 show that the performance indicators—compressor power requirement
and coefficient of performance—both improve as the inlet condenser water temperature decreases from 38°C to 16°
C. The required power consumption decreases and the coefficient of performance increases. These are expected
trends since lower condenser water temperatures provide a higher temperature difference between heat transfer
fluids resulting in increased cooling potential in the condenser. The refrigerating efficiency in Figure 7.10
decreases with decreasing condenser water temperature; however, this is expected since the system is designed for
optimal performance at an entering condenser water temperature of 31.4 °C. As observed in these figures, HFC-
236ea is predicted to outperform CFC-114 over a range of inlet condenser water temperatures. This is partly due
to the fact that measured heat transfer coefficients for HFC-236ea were found to be greater than those of CFC-114
[22],
0.35
0.3
0.25
>.
o
c
JD
O
!t 02
UJ
cn
c
5
ro
<5 0.15
o>
*-
a:
0.1
0.05
0
20	22	24	26	28	30	32	34	36	38
Entering Condenser Water Temperature, °C
Figure 7.10: Dependence of refrigerating efficiency on entering condensing water temperature
CFC-114 26fpi
HFC-236ea 26fpi
41
-------
Figure 7.1 i shows the refrigerant's saturation temperature in the condenser relative to the temperature of
the cooling water entering the condenser. The predicted saturation temperatures for CFC-114 and HFC-236ea are
nearly identical. The measured values of the condenser saturation temperature for CFC-114 agree with the
predicted values while the HFC-236ea data show the same trend but are generally higher than predicted values.
50
4b
O «
«
3 35
cs
k—
0)
Q_
E 30
ID
I—
C
2 25
m
D
W 20
0>
(/)
5 is
T3
C
o
° 10
Aa"

A ^
A
i *¦
A

A A A
A	..-A '
CFC-114 26f pi
	HFC-236ea 26fpi
A HFC-236ea data
a CFC-114 data
_j	i	i_
15
20	25	30
Entering Condenser Water Temperature, °C
36
40
Figure 7.11: Dependence of condenser temperature on entering condenser water temperature
Figure 7.12 shows the saturation temperature of the refrigerant in the evaporator as a function of the
entering condenser water temperature. The predicted saturation temperature for HFC-236ea is higher than the
predicted value for CFC-114 over the range of entering condenser water temperatures. The HFC-236ea data
compare well with predicted values while there appears to be less agreement between measured and modeled
values for CFC-114.
42
-------
CFC-114 26fp
HFC-236ea 26fpi
a HFC-236ea data
~ CFC-114 data
25	30
Entering Condenser Water Temperature, °C
Figure 7.12: Dependence of evaporator temperature on entering condenser water temperature
Figure 7.13 shows the relationship between the evaporator capacity arid the temperature of the water
entering the condenser. Since the evaporator capacity is fixed by holding the water-side conditions constant, the
predicted values for IIFC-236ea and CFC-114 are identical. Scatter is shown for measured values of HFC-236ea
while measured values of CFC-114 agree with predicted values In the model, the capacity is fixed for both CFC-
114 and HFC-236ea by the chilled water conditions.
Figure 7.14 shows the condenser capacity as a function of the entering condenser water temperature.
Predicted values for CFC-114 and HFC-236ea are nearly equal. Both the CFC-114 and HFC-236ea measured
values are higher than predicted values with significant scatter observed in the HFC-236ea data. This result is
consistent with previous results and discussion.
43
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CFC-114 26fpl
	HFC-236ea 26fpl
A HFC 2&6ea data
~ CFC-114 data
15
20	25	30
Entering Condenser Water Temperafure, °C
35
Figure 7.13: Dependence of evaporator capacity on entering condenser water temperature
400
	CFC-1 14 26fpl
	HFC-236ea 2fifpi
A HFC.236ea data
d CFC-114 data
15	20	25	30	35	40
Entering Condenser Water Temperature, °C
Figure 7.14: Dependence of condenser capacity on entering condenser water temperature
44
-------
Figure 7.15 shows the relationship between the refrigerant mass flow rate and the temperature of the
water entering the condenser. The measured and predicted values for CFC-114 are in close agreement while there
is significant scatter in the data for HFC-236ea As condensing water temperature increases, this figure shows an
increasing trend in the refrigerant mass flow rate. This is an expected trend since Figure 7.12 shows the
evaporator saturation temperature (and thus pressure) to be constant, and Figure 7.11 shows the condenser pressure
and temperature of the refrigerant to increase with increasing condenser water temperature It follows that the
enthalpy difference across the evaporator will decrease which requires an increase in refrigerant mass flow rate to
meet the given (or fixed) evaporator load.
4.5
AA
o) 3 c
- tr *"*
— — ** — —
"l 2.5
	CFC-114 26fpi
	HFC-236ea 26fpi
a HFC-236ea data
o CFC-114 data
05
15	20	25	30	35	40
Entering Condenser Water Temperature, °C
Figure 7.15: Dependence of refrigerant mass flow rate on entering condenser water temperature
Entering Evaporator Water Temperature
In this situation, the evaporator load is defined by a constant chilled water flow rate of 28.4 liters per
second, a chilled water inlet temperature of 7 °C, and a chilled water exit temperature ranging from 9.2 to 12.6 °C.
Additionally, the temperature of the water entering the condenser is held constant at 31.4 °C and the flow rate of
45
-------
the condenser water is held constant at 31.5 liters per second. As the cooling load is systematically varied, its
effect on various performance indicators may be observed.
As the temperature of the chilled water leaving the load and entering the evaporator increases while other
design operating conditions remain constant, there is an increasing (rend in the power required to drive the
compressor as shown by Figure 7.16. This is an expected trend because as the water temperature entering the
evaporator increases, the demand is increased on the evaporator. In order to accommodate this increased demand,
either the refrigerant mass flow rate or the enthalpy difference in the evaporator must increase in order to provide
enough heat transfer to maintain a constant chilled water exit temperature. The result is the need for more power
required to drive the compressor. The comparison of HFC-236ea and CFC-114 in Figure 7.16 shows that for the
range of chilled water temperatures entering the evaporator, HFC-236ea always requires less compressor power
when modeled as the working fluid
250
200
v 150
5
o
CL
O
V)
£/>
o
D- 100
E
o
O
50
- CFC-114 29pi
HFC-236ea 26fpi
10	11	12
Entering Evaporator Water Temperature, °C
13
14
Figure 7.16: Dependence of compressor power for CFC-114 and HFC 236ea on entering evaporator water
temperature
46
-------
A comparison of the coefficient of performance as a function of chilled water temperature entering the
evaporator shows that as the temperature increases the coefficient of" performance decreases. This means that as
the temperature increases., the increase of power required by the compressor is greater than the increase in cooling
capacity. Additionally, as shown in Figure 7.17, the coefficient of performance for HFC-236ea is higher than the
coefficient of performance for CFC-114 for the range of temperatures modeled.
Finally. Figure 7.18 shows that as the temperature of the chilled water entering the evaporator increases,
the refrigerating efficiency decreases. This figure also shows that only a narrow range of temperatures are both
realistic and optimum By definition, the refrigerating efficiency lies between the values of zero and one. There is
a general drop-in efficiency of 10 percent for a 2°C temperature increase. Thus, while the refrigerating efficiency
remained relatively stable for a wide range of condenser water temperatures, it is more sensitive to a change in
chilled water temperatures which essentially represent a change in capacity. HFC-236ea maintains a 3 to 5 percent
higher efficiency than CFC-114 for the range of temperatures modeled.
	CFC-114 2Qfpi
	HFC-236ea 26fpi
10	11	12
Entering Evaporator Water Temperature, °C
13
14
Figure 7.17: Dependence of refrigerating performance on entering evaporator water temperature
47
-------
0.45
0.4
0.35
>• 0.3
c
«
1	0.25
o>
c
2	0.2
-------
which are a function of average water temperature. Since the results are similar to the previous figures, the
following discussion will be brief.
Figure 7.19 is a plot of the chilled water temperature leaving the evaporator and the resulting effect on the
power consumption of the compressor As the temperature increases, the cooling load decreases and the
subsequent power required of the compressor diminishes. This plot also shows that for the range of temperatures
simulated HFC-236ea requires less predicted compressor power than CFC-114.
140
120
100
<:
-------
6
5

Q_
o 3
C
£
*o
<4—
o 2
o
1
0
6	6.5	7	7.5	8	8.5	9
Leaving Evaporator Water Temperature, °C
Figure 7.20: Dependence of refrigerating performance on leaving evaporator water temperature
The refrigerating efficiency varies as a function of the chilled water temperature exiting the evaporator as
illustrated by Figure 7.21. As the temperature increases, the refrigerating efficiency increases. Raising the
temperature of the set point reduces the water temperature difference across the evaporator and thereby reduces the
evaporator cooling capacity. Apparently, as the load is reduced, the corresponding compressor work is even less,
thereby causing the coefficient of performance to improve in relation to the maximum possible coefficient of
performance.
The trends of compressor power, coefficient of performance, and refrigerating efficiency as leaving
evaporator water temperature is varied over a range of 6.4 °C to S.7 CC are shown in Figures 7.19 through 7 21
The design point for the leaving chilled water evaporator temperature is 7 °C. As this value increases, the shaft
power requirement decreases, and the coefficient of performance and refrigerating efficiency increase.
50
-------
0.45
0.4
0.35
g- 0.3
c
¦V
o
£ 0.25
D)
C
? 0.2
-------
-ChC-114 26!p)
• HFC -236
-------
-CFC-1 14 26fpl
• HFC-236ea 26fpl
24
25
26	27	28	29
Evaporator Water Flow Rate, l/s
30
31
32
Figure 7.24: Dependence of refrigerating efficiency on evaporator water flow rate
Finally, as the evaporator water flow rate increases, Figure 7.24 illustrates that the refrigerating efficiency
steadily decreases, and again, HFC-236 is predicted to outperform CFC-114 for the range simulated.
The flow rate of the evaporator chilled water is varied to see its effect on system performance. Figures
7.22 through 7.24 show the significance of varying these parameters. As with the previous results, HFC-236ea is
predicted to outperform CFC-114 in every test case. The evaporator flow rate has a more significant effect for each
liter per second than does the condenser water flow rate as will be shown next. As the evaporator chilled water
flow rate increases, the load on the evaporator also increases It follows that the corresponding shaft power
required increases while the other performance indicators decrease.
Condenser Water Flow rate
As the condenser sea water flow rate is increased, the rate of heat transfer from the condenser increases
and the result is a decrease in power requirement and increases in coefficient of performance and refrigerating
efficiency. These observations are illustrated by Figures 7.26 through 7.28, respectively. In each case the change
is small over the range of condenser water flow rates simulated Additionally, in each case, the model predicts that
HFC-236ea will perform significantly better than CFC-114 under the same operating conditions.
53
-------
120
100
8C
o
n.
CL
£
o
O
60
40
20
-CFC-1 W 26fpl
HFC-236ea 26fpJ
23
29
30	31	32	33
Condenser Water Flow Rate, l/s
34
35
4.5
G> 3 5
o
t
-------
0.35
	CFC-1-U 26fpi
	H f C -236e a 20fpl
28	29	30	31	32	33	34	35
Condenser Watei Flow Rate, I/i>
Figure 7.27: Dependence of refrigerating efficiency oil condenser water flow rate.
SUMMARY
A parametric study was conducted using the computer program developed. Five parameters, including
entering condenser water temperature, entering and leaving evaporator water temperatures. 3nd condenser and
evaporator water flow rates, were tested over an appropriate range and comparison plots were generated. The
results of the model at design conditions suggest that HFC-236ea would outperform CFC-114 in a 440-kilowatt
centrifugal chiller. At design conditions, identified in Table 6.1. the coefficient of performance was modeled to be
12.9 percent greater for HFC-236ea than for CFC-114. The predicted shaft power required is 11.3 percent less for
HFC-236ea, and the refrigerating efficiency is predicted to be 12.9 percent higher for HFC-236ea.
The results of this study differ from those presented by Bare [7], The difference can be attributed in part
lo the effects of heal transfer which were taken into account in this study by incorporating correlations for
measured heat transfer coefficients for the evaporator and condenser. The measurements were taken in a separate
study using a single-tube heat transfer test facility at Iowa State University [22], The reported results present pool
boiling and condensation heat transfer coefficients as functions of heat flux at constant saturation temperature for
10.23 fins per centimeter tubes and two refrigerants. The results showed that for a given heat flux and constant
saturation temperature for a single 10.23 fins per centimeter tube, HFC-236ea performed with slightly higher heat
transfer coefficients than CFC-114. This helps to explain some of the trends seen in the previous figures and is
responsible, in part, for the predicted higher performance of HFC-236ea as a drop-in substitute for CFC-114.
0.3
0.25
UJ
CO
0.15
0.1
0.05
55
-------
CHAPTER 8. REFERENCES
[1]	Molina, M.J. and F.S. Rowland. "Stratospheric sink for chlorofluoromethanes: chlorine atom-catalysed
destruction of ozone." Nature 249 (1974): 810-814.
[2]	Wuebbles, D.J. "The role of refrigerants in climate change." International Journal of Refrigeration 17
(1994): 7-17.
[3]	Budiansky, S. "The doomsday myths." U.S. News & World Report (December 1993): 81-91.
[4]	Doyle, T.J., W.K Raymond, and A.L. Smookler. "Surface ship machinery - a survey of propulsion,
electrical and auxiliary system development." Marine Technology 29(3) (1992): 115-143.
[5]	Miro, C.R., and J.E. Cox. "Global environment policies enter critical phase in the near-future." ASIIRAE
Journal (March 1994): 14-15.
[6]	Helmick, Richard L., Bruce G. Unkel, Robert A. Cromis, and A. Lynn Hershey. "Development of an
advanced air conditioning plant for DDG-51 class ships," Naval Engineers Journal (May 1987): 112-123.
[7]	Bare, J.C. "Simulation of performance of chlorine-free fluorinated ethers and fluorinated hydrocarbons to
replace CFC-11 and CFC-114 in chillers." ASIIRAE Transactions 99 (1993): 397-407.
18] Threlkeld, J.L. Thermal Environmental Engineering, 2nd Ed., Englewood Cliffs, NJ: Prentice-Hall, 1970.
[9]	Smith, T.E.. R.M. Nelson, and M.B. Pate. "An interactive computer program for analyzing refrigeration
cycles in hvac courses." ASHRAE Transactions 93 (1987): 870-882.
[10]	Braun, J.E., J W. Mitchell, S. A. Clein, and W.A. Bechman. "Models for variable-speed centrifugal chillers."
ASHR4E Transactions 93 (1987): 1794-1813.
[11]	Jackson, W.L , F.C. Chen, and B.C. Hwang. 'The simulation and performance of a centrifugal chiller."
ASIIRAE Transactions 93 (1987): 1751-1707.
[12]	Wong. S.P.W., and S.K. Wang. "System simulation of the performance of a centrifugal chiller using a shell-
and-tube-type water-cooled condenser and R-l 1 as refrigerant." ASHRAE Transactions 95 (1989): 445-454.
[13]	Domanski, P.A., and M.O. McLinden. "A simplified cycle simulation model for the performance rating of
refrigerants and refrigerant mixtures." International Journal of Refrigeration 15(1992): 81-88.
[14]	Gauger, D C, H.N. Shapiro, and M.B. Pate. "Alternative Technologies for Refrigeration and Air-
Conditioning Applications. EPA-600/R-95-066 (NTIS PB95-224531), (May 1995), U.S. Environmental
Protection Agency, Research Triangle Park, NC.
[15]	Stoecker, W.F. Refrigeration and Air Conditioning, 2nd Ed, New York: McGraw-Hill, 1982.
[16]	Langley, B.C. Refrigeration and Air Conditioning, 3rd Ed, Englewood Cliffs, NJ. Prentice-Hall, 1986.
[17]	Schultz, J.M. 'The polytropic analysis of centrifugal compressors." Journal of Engineering for Power:
Transactions of the ASME (1962): 69-82.
56
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118] National Institute of Standards and Technology. Thermodynamic Properties of Refrigerants and
Refrigerant Mixtures Database (REFPROP version 4.01). Gaithersburg, MD, 1990.
119] Morrison, G., and M. O McLinden. "Application of a hard sphere equation of state to refrigerants and
refrigerant mixtures ." National Bureau of Standards Technical Note 1226, National Bureau of Standards.
Department of Commerce, Washington, DC, 1986
[20]	Moran. M.J , and H.S. Shapiro Fundamentals of Engineering Thermodynamics, 2nd Ed., New York John
Wiley and Sons, 1992
[21]	Incropera, F.P., and D P. DeWitt. Fundamentals of Heat and Mass Transfer, 3rd Ed.. New York: John
Wiley and Sons, 1990.
[22]	Huebscli, W Heat transfer evaluation of R-236ea and R-IN in condensation and evaporation. Master's
thesis, Iowa State University, Ames, 1A, 1994
|23] Stoecker. W F Design of Thermal Systems, 3rd Ed. New York: McGraw-Hill, 1989
57
-------
APPENDIX A. COMPUTER PROGRAM
PROGRAM CHTLLER
IMPLICIT DOUBLE PRECISION (A-H,0-Z)
LOGICAL LBUB, [.CRIT, LL1QI, LVCON
C "PREFS" IS USED TO PASS THE VALUE OF LEQN
COMMON /PREFS/ NUNITS,NREFST,INTACT,ICLMN,JCLKiN(7),IEQN
C "ESDATA" CONTAINS THE VALUES OF THE MOLECULAR WEIGHTS IN CRIT(1,N)
COMMON /ESDATA/ COEFF(10,40),CRIT(5,40)
C "MOLX" IS USED TO PASS THE VALUES OF THE MOLAR COMPOSITION IN X
COMMON /MOLX/ WMI(5),WMR,WMX,X(5)
C COMMON BLOCKS "CMNOM" AND "HREFI" GIVE ACCESS TO THE NAMES OF THE
C COMPONENTS: LNAME IS THE CHEMICAL NAME AND HREF IS THE REFRIGERANT #
CHARACTER*30 LNAME
CHARACTER* 10 SYNM
C,HARACTER*6 HREF
COMMON /CMNOM/ SYNM(5),LNAME(5)
COMMON /HREF1/ HREF(0:40)
DIMENSION XL(5),XV(5)
C dimension F(5,5), FT(5,5),XW(5)! needed with mixtures
common/group 1 /Tee.T el,Tce,Qe,TcI,Qc, V2,p2,uae,uac
comnion/evap/hiev,atevsi,atevso,tt\vev,ebc
common/cond/hicd,atcdsi,atcdsc\U\vcd,dtcdi,dtcdo,cdc
common/misc/eLMTD,cLMTD,itube,ir(5)
double precision ntev.ntcd
C These statements are for use with L1NPACK subroutines
real a(50,50),b(50),z(50)
dimension y(3),yold(3)
integer ipvt(50),lda
data lda,n/50,3/
data y,'40,100,110/ ! initial estimates: y(l)~Te, y(2)^Tc, y(3)=iT2
data tol/.001/	! convergence criteria for N-R iteration
C Design conditions:
data Tee,Tee,Tel,gpme,gpmc/88.,50.67,44.,450.,500./
data pexp,dsh,dsc/1.05,5.,5./
data PI/3 1415927/
C Heat Exchanger data for Laboratory 125-ton AC plant
data XLTEV/7.228/,NTEV/246./,XLTCD/7.228/,NTCD/246./
C
C FOR PURE FLUID CALCULATIONS, IT IS RECOMMENDED TO USE
C IEQN - 2 UNLESS THE FLUTD IS NH3 (FOR NH3 USE IEQN = 3)
C THIS APPLIES THE MODIFIED BENEDICT WEBB RUBIN EOS IF AVAILABLE,
C OTHERWISE THE EXTENDED CORRESPONDING STATES MODEL
C
C INITIALIZE REFPROP VARIABLES
58
-------
NC = .1	! number of components
print*,'input refrigerant 11 = R114, 29 = R236ea'
read*,ir( 1)	! component name
X( 1) = 1 .DO	! mole fraction of component
IEQN = 2	! ECS model
C CHOOSE REFERENCE STATE FOR ENTHALPY AND ENTROPY: NREFST
C 1: H,S = 0 FOR LIQUID AT NBPT
C 2: H.S = 0 FOR LIQUID AT -40C (ASHRAE)
C 3: H = 200KJ/KG, S=1 KJ/KG K FOR SATURATED LIQUID AT OC (HR)
C
NREFST = 2
C
CALL BCONST(NC,IR) ! To obtain EOS parameters from BLOCK DATA
C WRITE NAME OF CHOSEN FLUID
WRITE(* *) IR(1)/ ',HREF(IR(1)),' \LNAME(1)
C
C set up an output file
opcn(unit=13, file='results')
! Plain lube geometry
! (root diameter - 2*wall thickness) inches
! (root diameter + 2*fin height) inches
! 26fins per inch tube geometry
2 print*,'input tube l)Plain 2)26fins per inch 3)40fins per inch 4)TurboB'
read*,itube
if(itube.gt.4)goto 2
if(itube.eq. l)then
DTE VI=.017312*39.3696
DTEVO=.019446*39.3696
DTCDI-017312*39.3696
DT CDO=. 019446*39.3696
elseif(itube.eq.2)then
DTEVI-0143002*39.3696
DTE VO= 01905*39.3696
DTCDI=. 0143002*39.3696
DTCDO=.01905*39 3696
elseif(iajbe.eq.3)then
DTEVI =0155702*39.3696
DTEVO=.0188722*39.3696
DTCDI-0155702*39.3696
DTCDO- 0188722*39.3696
e!seif(itube.eq.4)then
DTE\T—.0160528*39.3696
DTEVO=.0184912*39.3696
DTCDI- .0155702*39.3696
DTCDCKO i 88722*39.3696
endif
! 40fins per inch tube geometry
! Turbo-B tube geometry (40fins per inch for condenser)
I print*,'input variable l)tce 2)tee 3)tel 4)we 5)wc'
read*,ivar
if(ivar.gt. 5)goto 1
print*,'input range (from,to,step)'
59
-------
read*jl j2j3
do 100 var=jl J2J3 ! parameter to vary
if(ivar.eq. I)then
tce=var
elseiffivar. eq. 2)lhen
tee=var
elseif(ivar.eq. 3)thon
tel=var
elseif(ivar.eq.4)then
we=var
elseif(ivar.cq.5)then
wc=var
endif
iter-1
C	
C	THERMOD YNAMIC ANALYSIS
C note; Be extremely careful with units!
C property subroutines (REFPROP) have units in molar SI:
C' T(K), P(kPa), v(L/mol), h(J/mol), etc...
C properties at the compressor inlet (statepoint 1)
20 T=(y(l)-459.67)/1.8	! (K)
C Use the BUBLT Routine to find saturation boundary
LBUB ^ .TRUE.
XL(1) = X(l)
CALL BlJBLT(T,XL.XV,P,VL,Yl ,LBUB,LCRIT)
if(lcrit)print*,'input above critical point'
pl=p/6.894S	! evaporator pressure (psia)
C GET OTILER SATURATED VAPOR PROPERTIES
if(dsh.eq.0.d0)then
CALL HCVCPS(1 ,T,V1 ,XV,HV,d;d,d)
elseif(dsh.gt.0.d0)then
THy(l )+dsh 1459.67)11.8	! (K)
lliqi=.false.
call vit(T,P,d.d,Vl ,LLrOI,LVCON)
if(lvcon)print*,'vit did not converge'
call hcvcps( 1 fT,Vl .XV,HV,d,d,d)
endif
111— hv/crit(I,IR(l))/2 326	! (Bta'lbm)
C properties at the condenser outlet (slate point 3)
T=(y(2)+459.67)/1.8	! (K)
C Use the BUBLT Routine to find saturation boundary
60
-------
CALL BUBLT(T,XL,XV,P,VL,VV,LBUB,LCR1T)
if(lcril)print*,'input above critical point'
p2—p/6.8948	! condenser pressure (psia)
C calculate enthalpy at the saturated vapor point
call hcvcps(l,T,VV,XV,HV,d.d,d)
h2s—hv/crit( 1 ,IR( 1 ))/2 326	! (Btu/lbm)
C GET OTHER SATURATED LIQUID PROPERTIES
if(dsc.eq.O.)then
CALL HCVCPS(l,T,VL,XL,HL.d,d,d)
elseifldsc.gt.O. )then
T=(y(2)-dsc+459.67)/1.8	! (K)
lliqi~ true.
call vit(T,P,d,d,VL,LLIQl.LVCON)
if(lvcon)print*,'vit did not converge'
call hcvcps(l,T,\T_,XL)HL,d,d,d)
endif
h3=hl/crit(l,lR( 1 ))/2.326	1 (Btu/lbm)
C properties at the compressor outlet (state point 2)
C calculate v2 using compressor polytropic analysis
v2=(pl/p2V*(l./pexp)*vl	! (1/mol)
if(w.gt.v2)then
print*,'wet compression'
goto 100
endif
C isenthalpic expansion
h4=h3	! (Btu/lbm)
C estimate value of T2 (F)
T=(y(3);-459.67)/l.8	! (K)
C, calculate h2 = f(T2,v2)
CALL tlCVCPS(l,T,V2,XV)H>d,d,d)
li2 =li/crit( 1 ,IR( 1 ))/2.326	! (Btu/lbm)
C. Convert gpm to lbm/min
CALL WATER(tee, 1 ,d,d,d,dens,d)
We—gpme*35.314/264 17*denx
CALL WATER(tce,2,d,d,d;dens,d)
Wc-igpmc*35.314/264.17*dens
C
C — Average Evaporator Water Temp (Deg F)—
C
TWEVA=(teet tel)/2.
C
C — Evaporator Water Properties —
61
-------
c
CALL WATER(TWEVA,1,VSWEV,CNWEV,CPWEV,DNWEV,FFEV) ! fresh wa»er
C
Qe=we*cpwev*(tee-tel) I (kVV)
Wr ~Qe/(hl-h4) ! (lbni/inin)
Qcl=Wr*(h2-h2s)
Qc2 Wr*(h2s-h3)
Qc=Qcl+Qc2	! (kW)
C iterate to find condenser leaving wafer temperature, tcl
tcl=tce+5 ! estimate
do 210 i= 1,100
t=tcl
C — Average Condenser Water Temp —
C
TWCDA-(tce+tclV2
C
C — Condenser Water Properties —
C
CALL WATER(twcda,2,vs\vcd,cnwed,cpwcd,dnwcd,d) ! seawater
C
tcl- tce+qc/wc/cpvvcd
C
C — Check for Convergence —
C
lF(ABS(tcl-t).LT..001) GOTO 200
IF<7.EQ. 100) THEN
print*,' iNo convergence on tcl'
STOP
END IF
210 continue
C		-		
C	HEAT TRANSFER ANALYSIS
C	
C — ALL COEFFICIENTS BASED ON GIVEN TUBE DIAMETERS —
C
p	evaporator ***************
c
C — Total Evap Tube Outside Surface Area —
200 ATEVSCMNTEV*PI*DTEVO*XLTEV/l 2.
C
C — Evap Tube Inside Surface Area —
C
ATEVSI=NTEV*PI*DTEVI*XLTEV/12.
C
C — Evap Total Tube Flow Area —
C
ATEVF=NTE V*P1 *(DTEVI/12.)**2/8. ! (2 pass hx)
62
-------
c
C — Evap Tube Water Velocity —
C
CALL WATER(tee,l,VISC,COND,CP.DWEVS,FOULF) ! fresh water
vwevt=we/dwevs/atev#60.
C
C — Evap Tube Reynolds Number —
C
REEVT=DNWEV*VWEVT*DTEVI*3600./12./VSWEV
C
C — Evap Tube Prandtl Number —
C
PREVT=CPWEV*VSWEV/CNWEV
C
C. — Iteration for Tube Wall Temp and Water Heat Transfer CoefT—
C
C — Initial guess for Tube Wall Temperature —
C
TTWEV=v(l)
C
C — Iteration Loop —
C
DO 70 1=1,100
T=TTWEV
CALL WATER(TTWEV,l,VSEVTW,COND,CP,DENS,FOULF) ! freshwater
C
C — Water Side Coeff — (Btu/hr/sqft/F)
C
H1EV= 027*REEVT**.8*PREVT**.333*(VSWEV/VSEVTW)**.14
& *CN WEV/DTEV1* 12.
C
C — Tube Wall Temp —
C
TTWEV=(teei tel)/2.-Qe*60./HlEV/ATEVSl
C
C — Check for Convergence —
C
IF(ABS(TTWEV-T).LT..001) GOTO 80
IF(I.EQ. 100) THEN
print*,' No convergence on evap tube wall temp'
STOP
END IF
70 CONTINUE
C
£ **+*****+*++**+ CONDENSER ***************
C
C
C — Condenser Total Tube Flow Area (sqft) — ! (2 pass lux)
63
-------
c
80 ATCDF=NTCD*PI*(DTCDI/12.)**2/8.
C
C — Total Condenser Tube Outside Surface Area (sqft) —
C
ATCDSO=NTCD*PI*DTCDO*XLTCD/12.
C
C — Total Condenser Tube Inside Surface Area (sqft) —
C
ATCDSI = NTCD*PI*DTCDI/12.*XLTCD
C
C — Tube Water Velocity —
C
CALL WATER(tce,2,VTSC,COND,CP,DWCDS,FOULF)! seawater
v wcdt^wc/dwcds/atcdf/ 6 0.
C
C — Condenser Tube Reynolds Number —
C
RECDT = dnwcd* V WC DT* DT CD1*3600. /12./vs wed
C
C — Condenser Tube Prandl Number ~
C
PRCDT = cpwcd*vswcd/cnwcd
C
C — Initialize Tube Wall Temperature —
C
TTWCD - y(2)
C
C — Iteration Loop for Tube Wall Temp, and Water Heat Transfer CoefT
C
DO 110 J-1,100
T = TTWCD
CALL WATER(TTWCD,2,VSCDTW,COND,CP,DENS,FOULF)! seawater
C
C — Water Side Coefficient —
C
HICD = .027*RECDT**.8+PRCDT**.333+(VSWCD/VSCDTW)**.14
& *CN WCD* 12.,OTCDI
C
C — Tube Wall Temperature —
C
TTWCD (tce+Tcl)/2.+Qc*60VHTCD/ ATCDSI
C
C — Check for Convergence —
C
IF(ABS(TTWCD-T).LT..001) GO TO 120
IF (J ,EQ. 100) THEN
print*.' No convergence on cond tube wall temperature'
STOP
END IF
110 CONTINUE
64
-------
C	Newton-Raphson Iteration Scheme Ref: Stoecker
C Keep old values of variables to later test for convergence
120 do 25 l-l,n
25 yold{i)=y(i)
C Calculate magnitudes of the functions at the temporary values
C of the variables (B matrix)
do 30 i=l,n
30b(i)-ff(i;y)
C Calc magnitudes of partial derivatives of all functions with respect
C to all variables (A matrix)
do40i-l,n
do40j~i.il
40 a(ij)=df(ij,y)
C Call UNPACK subroutines to solve simultaneous equations
call sgeco(a,lda,n,ipvt,rcond,z)
call sgesl(a,lda,n,ipvt.b,0)
C Corrected values of the variables (y new — y old - (yt - yc))
do 35 i=J,n
35 v(i) y(i)-b(i)
C Check for convergence
iflagMD
do 50 i=l,n
50 if(abs(y(i)-yold(i)).gt.tol)iflag- 1
if(iflag ne.l)goto 90
iter iter+1
goto 20
C	
C	Final Calculations
90 call water(tee,l,d,d,d,dens,d)
gpme=we/35.314/dens*264.17
call water(tce,2,d,d,d,dens,d)
gpmc=wc/35.314/dens*264.17
copr=(hl-h4)/(h2-hl) ! COP refrigeration side
coph= 1)* 144/778.17
+*.0353/crit(l,ir(l))*1000./2.2 ! polytropic work (Btu/lbm)
effr~copr/copc	! refrigerating efficiency
pact=Wr*(h2-h 1)	! compressor power (actual)
65
-------
ue=uae/atevso*60.
uc—uac/atcdso* 60.
write( 13,1 OSOjvar.hicd.cdc^iac.hiev.ebc.uae
print*,'wr = ',wr
print*,'h 1,1)4 = ',h I ,h4
print*,'te =' ,y( 1)
write(*, 1030)var,hicd,cdc,uac,hiev,ebc,uae
1030 format( lx,f9.2,6fl 0.0)
C write( 13,10.10)var,y(2),y( l);y(3),clmtd,elmtd
C +,wr,qe,qc,pact,ppol,copr,copli,wp,e0p,effr
C 1010 format(lx,f9 2,5f7.2,5fl0.2,3f6.2,2f6.3)
C wrile(*,l 020)var,hiev,ebc,hicd,cdc,ue,uc
C 1020 foraiat(lx,5fl0.0,2x,2fl0.0)
C write(*, 1020)var,y(2),y( 1 ),y(3),clmtd,elmtd,copr,uae,uac
C 1020 format(lx,f9.2,5f7.2,f6.2,2f9.2)
100 continue
end
C These are the functions to be set equal to zero,
function ff(i,y)
implicit double precision(a-h,o-z)
logical lliqi,lvcon
COMMON /MOLX/ WMI(5),Wlvra,WMX\X(5)
common/group 1 /Tee,T el ,T ce,Qe,T cl,Qc, V2,p2,uae,uac
cominon/evap/hiev,atevsi,atevso,tt\vev,ebc
comnion'cond''hicd.atcdsi,atcdso,ttwcd,dtcdi,dtcdo,cdc
conimon'niisc/cLMTD,cLMTD,ilube,ir(5)
dimension y(50)
COMMON /ESDATA/' COEFF( 10,40),CRIT(5,40)
goto( 1,2,3)i
1 eLMTD=((tee-y( l))-(tel-y(l)))/log((tee-y(l))/(tel-y(l»)
C convert Qe to Heat Flux (qef) in SI units
qef qe/atevso*60. ! Btu/lir/fLA2
qe^qeCO. 3171/1000. ! k\V/mA2
C choose correlation based on tube type
C correlations based on data taken at the ISU H.T. Test Facility
if(itube.eq.2 .and. ir(l).eq. 1 l)then
C File E11426P.TXT
C Standard Deviation = 3.265132E-02
C Error Sum of Squares = 3 .198326E-03
C Coefficients	Significant Figures F test (1/3) Pr > F
C C0 = .8431786	X	28.69 0.012
C C 1 = .1359888	X+0.81	125.41 0.001
66
-------
C C2--8.738483E-04	X+0.22	16.04 0.025
ebc= .8431786 + . 1359888 * qef-8.738483E-04 * qef**2
elseiffitube.eq 3 and. ir(l).eq.ll)then
C File-E11440PTXT
C Standard Deviation = 2.006233E-02
C Error Suin of Squares 1.207491E-03
C Coefficients	Significant Figures F test (1/3) Pr > F
C CO : .5452108	X	34.17 0.009
C CI = 1244462	X+0.96	298.86 0.000
C C 2 ~ -7.701729E-04	X+0.36	35.52 0.009
ebc= .5452108 + .1244462 * qef-7.701729E-04 * qef**2
elseif(itube.eq.l .and. ir(l).eq.ll)then
C File = El 14PLP.TXT
C Standard Deviation - 1 865557E-02
C Error Sum of Squares = 1.044091E-03
C Coefficients	Significant Figures F test (1/3) Pr > F
C C 0= .5603604	X	39.64 0.007
C C 1 = 9.845413E-02	X-0.85	205.56 0.001
C C 2 = -7.094109E-04	X-10.31	33.16 0 010
ebc- .5603604 + 9.845413E-02 * qef-7.094109E-04 * qef**2
elseif(itube.eq.4 .and. ir(l).eq. 1 l)then
C File -Ell 4TBP.TXT
C Standard Deviation = .346298
C Error Sum of Squares ¦ .3597669
C Coefficients	Significant Figures F test (1/3) Pr > F
C CO--10.90721	X	43.00 0.007
C CI 2.131048	X +0 89	274.83 0.000
C C 2 = -3.048937E-02	X ! 0.65	174.02 0.001
ebc~-l 0.90721 + 2.131048 * qef-3.048937E-02 * qef**2
elseif(itube.eq.4 .and. ir(l).eq.29)then
C File = E236TBR0.TXT
C Standard Deviation = .1865692
C Error Sum of Squares = . 1044243
C Coefficients	Significant Figures F test (1/3) Pr > F
C C0 = - 3412447	X	0.15 >0.3
C C 1 — 838187	X+1.99	148.31 0.001
C C 2 = -1.197194E-02	X+1.74	92.40 0.002
ebc—.3412447 + 8381S7 * qef-1.197I94E-02 * qeP*2
clseif(itube.eq. 1 .and. ir(l).eq.29)then
C File = E236PLP0.TXT
C Standard Deviation — I 864I38E-02
C Error Sum of Squares = 1.042504E-03
C Coefficients	Significant Figures F test (1/3) Pr>F
C C 0 = 1.635299	X	339.18 0.000
C C 1 - 8.369637E-02	X-0.31	146 78 0.001
C C 2 = -7.727059E-04	X-0.12	38.43 0.008
67
-------
ebc- 1.635299 - 8.369637E-02 * qcf-7.727059E-04 * qef**2
e1seif(itube.eq.2 and. ir(l).eq.29)tlien
C File = E23626P0.TXT
C Standard Deviation - 1.816126E-02
C Error Sum of Squares - 9.89494E-04
C Coefficients	Significant Figures F test (1/3) Pr > F
C C 0 = 2.22797	X	690.21 0.000
C C 1 ~ .1742529	X+0.49	691.38 0.000
C C2--1 766886E-03	X+0.10	215.76 0.001
ebc 2.22797 + .1742529 * qef-1.76688GE-03 * qef*2
elseif(itube.eq.3 and. ir(l).eq.29)lhen
C File = E23640P0.TXT
C Standard Deviation = 1.033823E-02
C Error Sum of Squares - 3 206369E-04
C Coefficients	Significant Figures F test (1/3) Pr > F
C CO- 1411047	X	837.85 0.000
C CI = 1529635	X+0.64	1639.40 0.000
C C 2 -1.333973E-03	X+0.18	383.88 0.000
ebc= 1 411047 + .1529635 * qef-1 333973E-03 * qef**2
endif
C convert evaporator boiling coefficient (ebc) to english units
ebc-ebc* 1000. ! W/mA2/K
ebc=ebc*0.17612 ! Btu/hr/ftA2/F
UAe=(l./(hiev*atevsi)+l.Aebc*atevso))**(-l.)/60. ! kW/F
fl=Qe- U Ae * eLMT D
return
2cLMTD-{(y(2Vtce)-(y(2)-Tcl))/log((y(2)-tce)/(y(2)-Tcl))
C convert Qc to SI units
qcf=qc/atcdso*60. ! Btu/hr/ftA2
qcf~qc6'0.3171 /1000. ! kW/mA2
C choose correlation to use
if(itube.eq 2 and. ir(l).eq ll)then
C File = CI 1426P.TXT
C Standard Deviation - 4 739834E-02
C Error Sum of Squares = 6.739808E-03
C Coefficients	Significant Figures F test (1/3) Pr>F
C C 0 = 3.620498	X	251.58 0.000
C CI .1494268	X+0.22	72.88 0.003
C C 2 = -2.087891F.-03	X-0.03	44.58 0.006
68
-------
cdc= 3.620498 i- .1494268 * qcf-2.087891E-03 * qcf**2
elscif((itube.eq.3 or. itube.eq.4) and. ir(l).eq. 1 l)then
CFile-C11440P.TXT
C Standard Deviation — .1500359
C Error Sum of Squares = 6.753228E-02
C Coefficients	Significant Figures F test (1/3) Pr > F
C C 0= 2.967423	X	17.87 0.021
C CI- .2155206	X-rO.46	15.51 0.026
C C 2 = -2.98224E-03	X -0.20	9.04 0.050
cdc= 2.967423 + .2155206 * qcf-2.98224E-03 * qcf**2
elseif(itube.eq. 1 and. ir(l).eq.l l)then
C File = C114PLP.TXT
C Standard Deviation " 1.139951E-02
C Error Sum of Squares = 2.598976E-04
C Coefficients	Significant Figures F test (1/2) Pr>F
C C 0= 1.510632	X	398.69 0.002
C C 1 =-2.106263E-02	X-0.3I	11.11 0.069
C. C 2 3.159286E-04	X-0.59	6.48 0.111
cdc= 1.510632 -2.106263E-02 * qcf t- 3.159286E-04 * qcf**2
elseif(itube eq.2 .and. ir(l).eq 29)then
C File - C23626P.TXT
C Standard Deviation - 4.463284E-02
C Error Sum of Squares = 5.97627E-03
C Coefficients	Significant Figures F test (1/3) Pr > F
C C 0 = 4 093141	X	388.58 0.000
C CI =.1415326	X+0.14	77.74 0.003
C C2 = -l.620209E-03	X-0.20	31.53 0.010
cdc= 4.093141 ¦) .1415326 * qcf-1 620209E-03 * qcf**2
elseif((itube eq.3 .or. itube.eq 4) .and. ir(l).eq.29)then
C File = C23640P.TXT
C Standard Deviation = 3.473567E-02
C Error Sum of Squares = .0036197
C Coefficients	Significant Figures F test (1/3) Pr > F
C C 0 = 2.953098	X	339.66 0.000
C CI =.187457	X +0.41	227.99 0.001
C C2 = -l.830992E-03	X-0.00	67.14 0.004
cdc~ 2.953098 - .187457 * qcf-1.830992E-03 * qcf**2
elscif(itube.eq.l and. ir(l).eq.29)the»
C File = C236PLP.TXT
C Standard Deviation = 2.255796E-02
C Error Sum of Squares = 1.526584E-03
C Coefficients	Significant Figures F test (1/3) Pr>F
C C 0 = 2.438327	X	557.62 0.000
C C 1--5.335892E-02	X-0.05	44.93 0.006
C C 2 = 7.070792E-04	X-0.33	24.58 0.014
cdc- 2 438327 -5.335892E-02 * qcf + 7.070792E-04 * qcf*"2
69
-------
endif
C convert cdc to english units
cdc=cdc*1000 ! W/mA2/K
cdc cdc*0.17612 ! B«u/hr/ftA2/F
UAc^( 1 ./(hicd*atcdsi>r1 ,/(cdc*atcdso))* *(-1. )/60. ! (kW/F)
jfiNQc-UAc*cLMTD
return
3T '(i)1 dely
df=(gl -g2)/(2 *dely)
return
end
SUBROUTINE WATER(TEMP,kWATER,VISC,COND,CP,DENS,FOULF)
implicit double precision (a-h,o-z)
C Mechanical and thermal properties of fresh water or seawater as a
C function of temperature within the range (32 - 158 F)
C
C Adapted From Oak Ridge Heat Pump Program
£*#*#**#*************#***#**+	******** si:*#***************#**#
C
C — Inputs —
C
C TEMP Water temperature (F)
C kWATER Kind of water. 1 - Fresh Water, 2 - Seawater
C
70
-------
C — Outputs —
C
C V1SC Viscosity of water (Ibm/ft-hr)
C COND Thermal conductivity of water (Btu/hr-ft-F)
C CP Specific heat of water (Btu/lbm-F)
C DENS Density of water (Ibm/cuft)
C FOULF Thermal fouling factor of water (lir-sqft-F/Btu)
C
IF(kWATER HE.2) THEN
C
C — Fresh Water —
C
VISC=3600 *(.62111E-11*TEMP**4-.29839E-8*TEMP**3+
& ,55359E-6*TEMP**2-.50665E-4*TEMP+-.2345E-2)
COND—,18797E-5*TEMP**2t ,85742E-3*TEMPl .2953
CP-.13254E-9*TEMP**4-.65618E-7*TEMP**3+.12373E-4*TEMP**2-
& ,10208E-2*TEMPH.029
C DENS=62.366-.0 i 63*(TE'MP-59.)
C
C The following equation for density is from Robert P. Benedict's text,
C FUNDAMENTALS OF TEMPERATURE, PRESSURE, AND FLOW MEASUREMENTS, Equation 15.7
C
DENS=62.2523+.978476E-2*TEMP-.145E-3*TEMP**2+.217E-6*TEMP**3
FOLJLF= 00010
ELSE
C
C — Seawater —
C
ViSC=3600.*(.62595E-l 1*TEMP**4-.29332E-8*TEMP**31
& .53738E-6*TEMP**2-.49718E-4*TEMP+.2408E-2)
COND- -.19056E-5*TEMP**2+, 85917E-3*TEMP+. 2884
CP=.952~(TEMP-32.)*5.88E-5
DENS"64.043-.00668*(TEMP-59.)
FOULF=. 00025
END IF
RETURN
END
C
C 	Definition of Variables	
C
C EVAPORATOR TUBE GEOMETRY **
C DTEVI - tube inside diameter (root diameter - 2*wall thickness) inches
C DTEVO - tube outside diameter (root diameter - 2*fin height) inches
C XLTEV - effective tube length per pass (feet)
C NTEV - number of tubes
C ATEVSO - total evaporator tube outside area (based on diameter)
C
C ** CONDENSER TUBE GEOMETRY **
C DTCDI - tube inside diameter (root diameter - 2*wall thickness) inches
C DTCDO - tube outside diameter (root diameter + 2*fin height) inches
71
-------
C	XLTCD - effective tube length per pass (feet)
C	NTCD - number of tubes
C	ATCDSO - total condenser tube outside area (based on diameter)
C
C	** EVAPORATOR **
C	GPME - chiller water flow rate (gpm)
C	EBC - boiling coefficient (btu/hr-ft**2-F)
C
C	** CONDENSER **
C	GPMC - condenser water flow rate (gpm)
C	CDC - condensing coefficient (btu/hr-ft**2-F)
72
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APPENDIX B.
NSWC AC PLANT INSTRUMENTATION SCHEMATIC
CONDENSER
WATER IN
COMPRESSOR—y
discharge
HOT-GAS-
BYPASS
LINE
IGV
COMPRESSOR SUCTION
I MOTOR
WATER IN
EVAPORATOR • Ts. p,
WATER OUT
ORIFICE PIATE
73
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APPENDIX C. SAMPLE NSWC DATA
Binary File: C:/DATA/T1244D05.JUL
Run Title HFC 236 EA 490 LBS
Test Date-	07/05/1994 1244 hours
Average of 25 data samples
Chn Description
Avg Value
Max Value
Min Value
Unit
00
Open Channel




01
Condenser Vapor
52.4724
52.5286
52.4155
psia
02
Compressor Discharge
53.2188
53.4079
53.1011
psia
03
Condenser Liquid
53.2145
53.2945
53.1486
psia
04
Evaporator Vapor Pressure
12.7113
12.7294
12.6963
psia
05
Compressor suction
12.5454
12.5725
12.5166
psia
06
Evaporator Liquid Pressure
12.6089
12.6452
12.5680
psia
07
Open Channel

13.2390
13.2380

08
Compressor Suction Temp2
36 8633
36.9007
36.7931
deg F
09
Compressor Suction Tempi
37.5992
37.6317
37.5551
deg F
10
Oil Cooler Inlet Temp
67.4745
67.4880
674635
deg F
11
Compressor Oil Sump Temp
123.772
123.790
123.759
degF
12
Chiller Water Inlet Temp
51.3134
51.3667
51.2593
deg F
13
Chiller Water Outlet Temp
44.0546
44.0852
44.0240
deg F
14
Oil Cooler Outlet Temp
69.2839
69.3016
69.2619
degF
15
Evaporator Vapor Temp
37 1099
37 1322
37.0710
degF
16
Condenser Water Inlet Temp
86.0839
86.1472
86.0216
deg F
17
Condenser Liquid Temp
101 821
101.840
101.807
deg F
IB
Evaoporator Liquid Temp
37.0662
37.0991
37.0348
degF
19
Condenser Vapor Temp
109.754
109.768
109.728
degF
20
Compressor Discharge Temp
109.443
109.470
109.358
deg F
21
Condenser Water Outlet Temp 94.2320
94.2812
94.1923
deg F
22
Motor Torque
3062.56
3079.39
3051.52
in-lb
23
Motor Speed
3593.84
3598.70
3590.98
rpm
24
Inlet Guide Vane position
38.7936
38.8078
38.7844
degree
25
Hot Gas Valve Position
-.700720
.327406
-1.18651
degree
26
Power input to compressor
.231624
.451580
-.034150
kW
27
Evaporator Water Flow rate
450.493
454.933
446.935
gpm
28
Condenser Water Flow rate
501.856
504.121
499.604
gpm
29
Oil Cooler Water Flow rate
14.4711
14.7647
14.0511
gpm
30
Motor Temperature
184.602
184.940
184.360
degF
31
Guideline Reference Temp
1.82038
1.90300
1.77330
degF
32
Spare RTD
1 83802
1.89820
1.77460
degF
13 Open Channel
Note: A data point consists of taking the average of 155 samples over 10 minutes. The point is considered stable
if the difference between the minimum and maximum temperatures of the inlet and outlet evaporator chilled water
is within 0.1 °F.
74
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APPENDIX D. CHILLER INFORMATION PROVIDED BV THE NSWC
The air conditioning plant consists of an open single-stage centrifugal compressor-motor driven unit, and
a condenser-chiller shell package. The compressor is direct-driven through a torquemeter station. Compressor
impeller speed is increased through an internal compressor gear arrangement. The refrigerant system is designed
to use CFC-114.
Compressor capacity is controlled to maintain the desired water temperature and to prevent motor
overloading. Control is achieved by varying the position of vanes located in the compressor suction inlet The
vanes are moved by a pneumatic operator which automatically responds to a chilled water thermostat.

TECHNICAL DATA
System caoacitv:
125 tons refrig with following design conditions;
Chilled water circulation
450 gpm
Temperature entering
50.7 CF
Temperature leaving
44 °F
Condenser water flow
500 gpm
Temperature entering
88 °F
Refrigerant charge
5001b. CFC-114 (approx )
Driveline_unil:

Centrifugal compressor

Type
Single-stage open
Capacity
125 tons
Impeller speed
11,918 rpm
Shell unit:

Condenser

Size
22 inches OD x 93.5 inches length
Class
B
Tvpe
Shell-and-tube
Number of tubes
246
Tube type
0.75" nominal diameter, 26 fins per inch, copper, 0.049" wall
Cooling surface
1138 square feet
Water passes
2
Circulating water
500 gpm
Inlet temperature
88 °F
Outlet temperature
95.9 DF
Refrigerant pressure
50 psig
Condensing temperature
104.8 °F
Cooler

Size
32 inches OD x 93.5 inches length
Class
B
Type
Shell-and-tube (flooded)
Number of tubes
246
T ube type
0.75" nominal diameter, 26 fins per inch, copper, 0.049" wall
Cooling surface
1138 square feet
Water passes
2
Circulating water
450 gpm
Inlet temperature
50.7 °F
Outlet temperature
44 °F
75
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