NISTIR 4748
The Performance of Chlorine-Free Binary Zeotropic
Refrigerant Mixtures in a Heat Pump
Jurgen Pannock
David A. Didion
Building and Fire Research Laboratory
Gaithersburg, Maryland 20899
nist
United States Department of Commerce
Technology Administration
National Institute of Standards and Technology
-------
PB92-1H9814
"'ST*11«* U.S. DEPARTMENT OF COMMERCE
(REV. 3-90) NATIONAL INSTITUTE OF STANDARDS AND TECHNOLOGY
V .
BIBLIOGRAPHIC DATASHEET
1. PUBLICATION OR REPORT NUMBER
NISTIR 4748
2. PERFORMING ORGANIZATION REPORT NUMBER
3. PUBLICATION DATE
DECEMBER 1992
4. TITLE AND SUBTITLE
The Performance of Chlorine-Free Binary Zeotropic Refrigerant Mixtures in a Heat Pump
5. AUTHOR (S)
Jurgen Pannock, David A. Didion
6. PERFORMING OROANOATION (IF JOINT OR OTHER THAN NIST, SEE INSTRUCTIONS)
U.S. DEPARTMENT OF COMMERCE
NATIONAL INSTITUTE OF STANDARDS AND TECHNOLOGY
QAITHERSBURQ, MD 20899
7. CONTRACT/GRANT NUMBER
a TYPE OF REPORT AND PERIOD COVERED
B. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (STREET, CTIV, STATE, BP)
EPRI EPA
3412 Hillviev Avenue . Alexander Drive & Highway 54
P.O. Box 10412 an Research Triangle Park, NC 27711
Palo Alto, CA 94303 AJTST
10, SUPPLEMENTARY NOTES
11. ABSTRACT (A20O-WORD (M 1£SS FACTUAL SUMMARY OF MOST SIQNIFICAMT INFORMATION. IF DOCUMENTINCLUDES A SIGNIFICANT BtBUOGRAPHY OR
LITERATURE SURVEY, MENTION IT HERE.)
The phase-out of the currently used refrigerants during the next decade requires fast and
accurate methods to evaluate possible alternatives for the existing refrigerants. This report
investigates possible replacement refrigerants for R22, where the replacements are binary
zeotropic mixtures of the following hydrofluorocarbons (HFCs)r R23, R32, R125, R134a, and
R152a. The method, that was chosen, is based on three steps:
1) determining possible mixture components,
2) evaluating all fifteen possible mixtures using a simulation program developed by NIST and
determining the best performing mixtures,
3) evaluating the best performing mixtures in a NIST build test facility.
Following this path, two refrigerant mixtures, R32/R134a and R32/R152a were found to perform
better than R22 with respect to COP and volumetric capacity for certain composition ranges.
The used simulation model proved to be a very precise tool in finding possible replacement
fluids and their possible performance advantages.
The results give the confidence that this time saving combination of simulation and testing is
a very powerful engineering tool.
12. KEYWORDS (8 T013 ENTRIES; ALPHABETICAL ORDER; CAPITALIZE ONLY PROPER NAMES; AND SEPARATE KEY WORDS BY SEMICOLONS)
binary zeotropic mixtures; CO?; performance advantage; R22; R23; R32; R125; R134a; 1143a;
R152a; test facility; volumetric capacity
13. AVAILABILITY
X
UHUMITED
FOR OFFICIAL DISTRIBUTION. DO NOT RELEASE TO NATIONAL TECHNICAL INFORMATION SERVICE (NTIS).
ORDER FROM SUPERINTENDENT OF DOCUMENTS, U.S. GOVERNMENT PRINTING OFFICE.
WASHINGTON, DC 20402.
ORDER FROM NATIONAL TECHNICAL INFORMATION SERVICE (NTIS), SPRINOFIEID.VA 22W,
14. NUMBER OF PRINTED PACES
87
IS. PRICE
AOS
ELECTRONIC FORM
-------
NISTIR 4748
The Performance of Chlorine-Free Binary Zeotropic
Refrigerant Mixtures in a Heat Pump
Jurgen Pannock
David A.Didion
IAG DW 13934749
EPA-600-R-92-017
December 1991
Building and Fire Research Laboratory
National Institute of Standards and Technology
Gaithersburg, MD 20899
U.S. Department of Commerce
Rockwell A. Schnabel, Acting Secretary
Technology Administration
Robert M, White, Under Secretary for Technology
National Institute of Standards and Technology
John W. Lyons, Director
Prepared for:
Electric Power Research Institute
Powell A. Joyner, Ph.D.
Manager, Advanced Projects
Custom Devision, Residential Program
3412 Hillview Avenue
Palo Alto, CA 94303
Prepared for:
U.S. Environmental Protection Agency
Robert V. Hendriks, Project Officer
Air & Energy Engineering Research Laboratory
Research Triangle Park, NC 27711
-------
Abstract
The phaseout of the currently used refrigerants during the next
decade requires fast and accurate methods to evaluate possible
alternatives for the existing refrigerants. This report inves-
tigates possible replacement refrigerants for R22, where the
replacements are binary zeotropic mixtures of the following
hydrofluorocarbons (HFCs): R23, R32, R125, R134a, R143a and R152a.
The method that was chosen is based on three steps:
1) determining possible mixture components.
2) evaluating all fifteen possible mixtures, using a simulation
program developed by NIST and determining the best performing
mixtures based on the simulation results.
3) evaluating the best performing mixtures in a test facility
built at NIST.
Following this path, two refrigerant mixtures, R32/R134a and
R32/R152a, were found to perform better than R22 with respect to
COP and volumetric capacity for certain composition ranges using
counterflow heat exchangers.
The simulation model proved to be a very precise tool in finding
possible replacement fluids and their possible performance
advantages.
The results give confidence that this time saving combination of
simulating and testing is a very powerful engineering tool.
iii
-------
Acknowledgements
This study was jointly sponsored by the Residential Building
Equipment & Systems Division of the Electric Power Research
Institute, Palo Alto, CA, under the direction of Powell A. Joyner
and the Air and Energy Engineering Research Laboratory of the
United States Environmental Protection Agency, Research Triangle
Park, N.C under the direction of Robert V. Hendriks. The Authors
would like to thank P.A. Domanski, R.Hampson, M.A. Kedzerski,
G. Morrison, W.J. Mulroy of NIST and R. Radermacher of the
University of Maryland for their helpful interactions on several
occasions.
iv
-------
TABLE OF CONTENTS
Abstract ...... ...... . iii
Acknowledgements ...... iv
List of Figures viii
List of Tables . xii
Nomenclature . xiii
1. Introduction . . 1
2. Simulations 1
2.1 Description of Input Data and Output Data ...... 2
2.2 Approach to Choose the Input Variables ....... 4
2.2.1 Input Variables Describing the Refrigeration
System • ~»•»»..»»...»• ......4
2.2.2 Input Variables Describing the Operation
Conditions 5
2.2.3 Processing the Simulation Results ....... 7
3. Selection Mixture Components .... 7
4. Pure Refrigerants 8
4.1 General Remarks ..... 8
4.2 R23 ............. ..... 10
4.3 R32 10
4.4 R12 5 10
4.5 R143a 10
4.6 R134a ...•«»*.*»•»».*«*•••••• 11
4.7 R152a 11
v
-------
5. Results of the Simulation Study 12
5.1 General Remarks . , 12
5.2 The Mixtures and their Projected Performance ... 12
5.2.1 R23/R32 13
5.2.2 R23/R125 ............ 13
5.2.3 R23/R143a 14
5.2.4 R23/R134a 14
5.2.5 R23/R152a ...... 14
5.2.6 R32/R125 ....... 14
5.2.7 R32/R143a 15
5.2.8 R32/R134a ........... 15
5.2.9 R32/R152a 16
5.2.10 R125/R143a 17
5.2.11 R125/R134a 17
5.2.12 R125/R152a . 17
5.2.13 R143a/R134a 18
5.2.14 R143a/R152a 18
5.2.15 R134a/R152a ............ 18
6. Discussion of the Simulation Results 18
7. The Mini-Breadboard Heat Pump 20
7.1 General Remarks 20
7.2 System Design ............. 20
7.3 System Instrumentation ......... 22
vi
-------
8. Test Results . , . 25
8.1 General Remarks 25
8.2 High Temperature Cooling Test (test condition 1A) . 25
8.3 Low Temperature Cooling Test (test condition IB) . 26
8.4 High Temperature Heating Test (test condition 1C) . 27
8.5 Low Temperature Heating Test (test condition ID) . 28
8.6 Liquid Line Heat Exchange (test condition 1A-LLHX) 29
8.7 Test Conclusions 31
9. Comparability of Computer Study and Test Results .... 32
10. Conclusions 34
11. Further Research 35
12. References ........ ...... 36
APPENDIX A: Simulation Data 37
APPENDIX B: Test Data 42
APPENDIX C: Uncertainty Analysis ... 61
vii
-------
LIST OP FIGURES
Fig. 1: Key Point Locations of the Simulation Program CYCLE11
Fig. 2: Vapor Pressure Curves of Selected Refrigerants
[lg p / (1/T) scale] ......
Fig. 3: System Setup of the Mini-Breadboard Heat Pump . . .
Fig. 4: Instrumentation of the Mini-Breadboard Heat Pump
Fig. A1 : Relative COP of R32/R134a & R32/R152a vs. R32
content; high temperature cooling condition
(1A)
Fig. A2 : Relative vol. capacity of R32/R134a &
R32/Rl52a vs. R32 content; high temperature
cooling condition
Fig. A3 J Relative COP of R32/R134a & R32/R152a vs. R32
content; low temperature cooling condition
(1A) ...........a...........
Fig. A4 : Relative vol. capacity of R32/R134a &
R32/Rl52a vs. R32 content; low temperature
cooling condition ... . .
Fig. A5 ; Relative COP of R32/R134a & R32/R152a vs. R32
content; high temperature heating condition
Fig. A6 : Relative vol. capacity of R32/R134a &
R32/R152a vs. R32 content; high temperature
heating condition ... .....
Fig. A7 : Relative COP Of R32/R134a & R32/R152a vs. R32
content; low temperature heating condition
(1A) .•»•«•¦«.
Fig. A8 : Relative vol. capacity of R32/R134a &
R32/R!52a vs. R32 content; low temperature
heating condition .......
Fig. Bl.l: Compressor speed vs. R32 mass fraction;
high temperature cooling condition (1A) . . .
Fig. B1.2 ; Compressor speed vs. R32 mass fraction;
low temperature cooling condition (IB) ....
3
9
21
23
38
38
39
39
40
40
41
41
43
43
viii
-------
Fig. B1.3 : Compressor speed vs. R32 mass fraction;
high temperature heating condition (1C) ... 44
Fig. B1.4 : Compressor speed vs. R32 mass fraction;
low temperature heating condition (ID) .... 44
Fig. B1.5 : Compressor speed vs. R32 mass fraction;
high temperature cooling condition with LLHX
(1A-LLHX) 45
Fig. B2.1 : Cooling COP vs. R32 mass fraction;
high temperature cooling condition (1A) ... 45
Fig. B2.2 : Cooling COP vs. R32 mass fraction;
low temperature cooling condition (IB) .... 46
Fig. B2.3 : Heating COP vs. R32 mass fraction;
high temperature heating condition (1C) ... 46
Fig. B2.4 : Heating COP vs. R32 mass fraction;
low temperature heating condition (ID) .... 47
Fig. B2.5 : Cooling COP vs. R32 mass fraction;
high temperature cooling condition with LLHX
(1A-LLHX) ........ 47
Fig. B3.1 ; Relative cooling COP vs. R32 mass fraction;
high temperature cooling condition (1A) ... 48
Fig. B3.2 : Relative cooling COP vs. R32 mass fraction;
low temperature cooling condition (IB) .... 48
Fig. B3.3 : Relative heating COP vs. R32 mass fraction;
high temperature heating condition (1C) ... 49
Fig. B3.4 : Relative heating COP vs. R32 mass fraction;
low temperature heating condition (ID) .... 49
Fig. B3.5 : Relative cooling COP vs. R32 mass fraction;
high temperature cooling condition with LLHX
(1A-LLHX) 50
Fig. B4.1 : Suction pressure vs. R32 mass fraction;
high temperature cooling condition (1A) ... 50
Fig. B4.2 : Suction pressure vs. R32 mass fraction;
low temperature cooling condition (IB) .... 51
Fig. B4.3 : Suction pressure vs. R32 mass fraction;
high temperature heating condition (1C) ... 51
ix
-------
Pig. B4.4 : Suction pressure vs. R32 mass fraction?
low temperature heating condition (ID) .... 52
Fig. B4.5 : Suction pressure vs. R32 mass fraction;
high temperature cooling condition with LLHX
(1A-LLHX) 52
Fig. B5.1 : Discharge pressure vs. R32 mass fraction;
high temperature cooling condition (1A) ... 53
Fig. B5.2 ; Discharge pressure vs. R32 mass fraction;
low temperature cooling condition (IB) .... 53
Fig. B5.3 : Discharge pressure vs. R32 mass fraction;
high temperature heating condition (1C) ... 54
Fig. B5.4 ; Discharge pressure vs. R32 mass fraction;
low temperature heating condition (ID) .... 54
Fig. B5.5 ; Discharge pressure vs. R32 mass fraction;
high temperature cooling condition with LLHX
(1A-LLHX)
Fig. B6.1 : Suction temperature vs. R32 mass fraction;
high temperature cooling condition (1A)
Fig. B6.2 ; Suction temperature vs. R32 mass fraction;
low temperature cooling condition (IB) . .
Fig. B6.3 i Suction temperature vs. R32 mass fraction;
high temperature heating condition (1C)
Fig. B6.4 ; Suction temperature vs. R32 mass fraction;
low temperature heating condition (ID) . .
55
55
56
56
57
Fig. B6.5 : Suction temperature vs. R32 mass fraction;
high temperature cooling condition with LLHX
(1A-LLHX) 57
Fig. B7.1 ; Discharge temperature vs. R32 mass fraction;
high temperature cooling condition (1A) ... 58
Fig. B7.2 : Discharge temperature vs. R32 mass fraction;
low temperature cooling condition (IB) .... 58
Fig. B7.3 : Discharge temperature vs. R32 mass fraction;
high temperature heating condition (1C) ... 59
Fig. B7.4 : Discharge temperature vs. R32 mass fraction;
low temperature heating condition (ID) .... 59
x
-------
Fig. B7.5 : Discharge temperature vs. R32 mass fraction;
high temperature cooling condition with LLHX
(1A-LLHX) ..... 60
Fig. CI : Specific heat regression curve for the heat transfer
fluid (40 %-mass ethylene-glycol, 60 %-niass water) 64
Fig. C2 ; Mass flow meter calibration curve for evaporator heat
transfer fluid measurements .. 66
Fig. C3 : Mass flow meter calibration curve for condenser heat
transfer fluid measurements ..... . . 66
xi
-------
LIST OF TABLES
1. Heat Transfer Fluid Temperatures by ANSI/ASHRAE
Standard 116-1983 6
2. Heat Transfer Fluid Temperatures Used for the Computer
Simulations ..... . 7
xii
-------
Nomenclature
CFC - Chlorofluorocarbon
COP - Coefficient of Performance
GWP - Global Warming Potential
h - enthalpy (kJ/kg)
HCFC - Hydrochlorofluorocarbon
HFC - Hydrofluorocarbon
HTF - Heat Transfer Fluid
LLHX - Liquid Line Heat Exchanger
MBHP - Mini-Breadboard Heat Pump
ODP - Ozone Depletion Potential
p - density (kg/rr3)
v - specific volume (mA3/kg)
q - specific capacity (kJ/kg)
qvol - volumetric capacity (kJ/mA3)
v - specific volume (m"3/kg)
VFRC - Volumetric Flow Rate in the Condenser
VFRE - Volumetric Flow Rate in the Evaporator
SUBSCRIPTS
numbers - position of properties with respect to Fig. 1
xiii
-------
l. Introduction
The growing concern about the environmental compatibility of
currently used refrigerants requires the research of new refri-
gerants and different refrigerating methods. The incompatibility
of the current CFC and HCFC refrigerants with the Earth's ozone
layer and the high greenhouse warming potential (GWP) of most of
the currently used refrigerants form the justification for this
research project. In this study, ozone-safe refrigerants with
relatively low GWP are investigated as zeotropic mixtures for heat
pump applications. Until now, this field was dominated by the
refrigerant R22.
The refrigerants considered in this study are chemical deriva-
tives of methane and ethane. With the requirement that the Ozone-
Depletion Potential (ODP) of the substances considered had to be
zero, xt followed that they were not allowed to contaxn any
chlorine or bromine (chlorine and bromine are the main catalytic
substances destroying the ozone layer). These requirements,
together with the goal to find a suitable working fluid for heat
pump applications, resulted in the selection of the following
substances as possible mixture components: R23, R32, R125, R134a,
Rl43a, Rl52a. In order to evaluate their performance, the
refrigerants are compared to R22, which is the commonly used
refrigerant m resxdentxal heat pumps in the USA*
In the theoretical part of this research, suitable refrigerant
mixtures for the possible replacement of R22 were determined. This
was accomplished by means of the application of the NIST-developed
simulation program CYCL111 [1]. The results of this computer study
are compared and analyzed with respect to their performance data
such as COP, volumetric capacity, discharge temperatures, discharge
pressure, and suction pressure.
The experimental part of this research project consisted of
tests, using the theoretically best performing refrigerant mixtures
as working fluids, in the Mini-Breadboard Heat Pump (MBHP) that was
developed and built at NIST. Those tested binary zeotropic
mixtures are R32/R134a and R32/R152a.
2. Simulations
In order to screen an initially 15 possible refrigerant mixtures
on a theoretical basis over their whole composition range, a
simplified refrigeration cycle program, CYCLE11, was used. The
program simulates a vapor-compression cycle that takes into account
common deviations of the heat pump cycle from the theoretical
process. The operating conditions are prescribed in terms of the
temperatures of the external heat transfer fluids. The heat
exchangers (condenser and evaporator) are generalized by an average
1
-------
effective temperature difference between the heat transfer fluid
and the refrigerant. An isenthalpic expansion across the expansion
device is assumed. The program also includes a rudimentary
compressor model and a representation of a suction line liquid line
heat exchange. For this research, however, the suction line to
liquid line heat exchange feature was not implemented. Ther-
modynamic properties of the refrigerants are calculated with the
Carnahan-Starling-DeSantis equation-of-state [2]. The program does
not require or use refrigerant transport properties. In order to
further understand the assumptions made for the simulation model
used in this analysis, the authors invite the interested individual
to read reference [1].
The simulation was used to give a reasonable idea about which of
the considered refrigerant mixtures can replace R22 as a heat pump
working fluid. Therefore, it was sufficient to limit the number of
simulation runs by assuming many parameters to be constant for all
four considered operating conditions. These parameters are:
• compressor speed
• refrigerant pressure drops in the heat exchangers
• polytropic compressor efficiency
• refrigerant liquid subcooling (leaving the condenser)
• refrigerant vapor superheat (leaving the evaporator)
• compressor pressure drop and heat transfer parameters
The authors recognized that these parameters would not be
constant in the test apparatus and therefore the comparability of
test and simulation results would be limited. However, the purpose
of the simulations was to screen and compare a large number of
refrigerant mixtures on a consistent basis. This task was satis-
fied.
2.1 Description of Input Data and output Data
The program input and output data consist of two parts;
• working fluid specification
• cycle specification
The working fluid specification includes identification of the
refrigerants and their composition in the mixture.
The refrigerant cycle specifications are defined by the following
eleven refrigerant states, which correspond to key locations in a
real system (Fig. 1), and are calculated by the program:
- 1- Suction line outlet, inlet to the shell of the hermetic
compressor
2
-------
- 2- Refrigerant state in the cylinder before the compression
process
- 3- Refrigerant state in the cylinder after the compression
process
- 4- Compressor shell outlet, condenser inlet
- 5- Saturated vapor refrigerant state in the condenser
- 6- Saturated liquid refrigerant state in the condenser
- 7- Condenser exit, liquid line inlet
- 8- Liquid line exit, inlet to adiabatic expansion device
- 9- Expansion device outlet, evaporator inlet
-10- Saturated vapor refrigerant state in the evaporator
-11- Evaporator outlet, suction line inlet
I-4 - compressor
4-7 - condenser
7-8 - liquid line
8-9 - expansion device
9-11- evaporator
II-1-vapor line
IWTF
Condenser
LLHX
LLHX
Expansion
device
10
LLHX
Ccrrtretssor
11 bypass
—I HTF
Evaporator
HTF-Heat transfer Ud
LLHX - liquid fete heat mchanQsr
Entropy
Fig. l: Key Point Locations of the Simulation Program CYCLE11
The following list presents the input data that must be provided
to the simulation program:
• Polytropic compressor efficiency
• Heat transfer fluid temperature at the condenser inlet
• Heat transfer fluid temperature at the condenser outlet
• Heat transfer fluid temperature at the evaporator inlet
• Heat transfer fluid temperature at the evaporator outlet
• Average temperature difference in the evaporator
3
-------
• Average temperature difference in the condenser
• Pressure drop in the condenser
• Pressure drop in the evaporator
• Refrigerant subcooling leaving the condenser
• Refrigerant superheat leaving the evaporator
• Refrigerant temperature at the compressor inlet
• Ratio of compressor RPM to reference RPM
• Heat transfer coefficient for process 1-2 (Fig. 1)
• Heat transfer coefficient for process 3-4 (Fig. 1)
• Pressure drop coefficient for process 1-2 (Fig. 1)
• Pressure drop coefficient for process 3-4 (Fig. 1)
The specification of the refrigerant temperature at the compres-
sor inlet is optional and defines the usage of the liquid line to
suction line heat exchanger (the LLHX feature was not used for the
simulation study).
2.2 Approach to Choose the Input Variables
The information given in section 2.1 shows system dependent
parameters and operating condition dependent parameters. For the
definition of these parameters, two different approaches have been
used.
The system dependent parameters (section 2.2.1), which are mainly
compressor specifications were determined in an empirical way.
The operating condition depending parameters (section 2.2.2) were
specified by air-air heat pump specifications, although the actual
tests were performed on the liquid-liquid MBHP.
2.2.1 Input Variables Describing the Refrigeration System
In order to specify the input variables for the compressor (e.g.
the heat transfer coefficients, the pressure drop coefficients, and
the polytropic compressor efficiency), the computer model was used
to generate output data over a range of different compressor speeds
at a certain operating condition using refrigerant R22. Then the
same heat transfer fluid conditions were used to run laboratory
tests with the MBHP, also using R22 as the working fluid. The
results of these laboratory tests were then compared with the
simulation results. By changing the above specified parameters in
the simulation program, the model was adjusted to reflect the real
system.
The goal of this procedure was to show the tendency of changes of
the coefficient of performance (COP) and of the volumetric capacity
with respect to different compressor speeds. It was not expected
nor intended to actually achieve the same absolute values in the
computer study as the measured data, since the simulation program
4
-------
employs several simplifications compared to a real system. This
means that the results of the computer study had to be evaluated on
a relative basis. They also had to be confirmed by laboratory tests
on the same relative basis. Nevertheless, they reflect the
relative system performance regarding the different refrigerants
and mixtures accurately.
2.2.2 Input variables Describing the Operation Conditions
The ASHRAE Standard ANSI/ASHRAE 116-1983 [3], which is based on
the DOE requirements, is used for the definition of the heat
transfer fluid temperatures. These test standards specify wet bulb
and dry bulb temperatures. Since the test rig is a water/glycol to
water/glycol heat pump system, the external heat transfer fluid
temperatures are specified by the temperature corresponding to the
dry bulb temperature of the air-air standard. For the heating test
conditions, the two steady state conditions for a single speed
compressor were used. These temperatures are given in table 1.
The outlet temperatures of the HTF appearing in that table are
calculated according to the following assumptions :
- the evaporator airflow rate is 187.90 m3/kWh
(400 scfm/ton) of refrigeration
the condenser airflow rate is 375.79 m3/kWh
(800 scfm/ton) of refrigeration (i.e. per 4.57197 kw
(1.3 tons) of rejected heat, assuming a COP of 3.0 for
the cycle performance)
Applying these assumptions, the calculated temperature change for
a 3.5169 kW (1 ton) evaporator load per 0.18878 m3/s (400 scfm)
airflow rate, is about 12.3 K (22 °F) . For a condenser load of
4.57197 kW (15,600 Btu/h) per 0.37756 m3 (800 scfm) airflow rate,
5
-------
Test A1
Test B2
Test C3
Test D4
°c
°C
°C
°C
(°F)
(°F)
(°F)
(°F)
Condenser Inlet
35.0
27.8
21.1
21.1
(95)
(82)
(70)
(70)
Condenser Outlet
43.2
VFRC5
VFRE6
VFRE6
(110)
Evaporator Inlet
26.7
2S.7
8.3
-8.3
(80)
(80)
(47)
(17)
Evaporator Outlet
14.4
VFRE6
VFRC5
VFRC5
(58)
TABLE l: Heat Transfer Fluid Temperatures by AHSX/ASHRAE
Standard 116-1983 £3]
the temperature change is about 8.2 °C (15 °F). This method
provides a set of inlet and outlet temperatures of the heat
transfer fluid for the steady state air conditioning test at 35 °C
(95 °F) condenser inlet temperature (test A). For the heating
conditions (Table 1, Test C & D), it was assumed that the airflow
rates (i.e. the liquid flow rates in the test rig) are not changed
in the indoor and outdoor part of the unit. Since any unit
experiences a change in performance under such a constraint, it is
not possible to accurately predetermine the outlet temperatures of
the HTF. In order to use reasonable values for those temperatures,
experimental tests with R22 were run at equivalent test A con-
ditions. The established liquid flow rates for test A were then
used to run experiments under the other three conditions. The
heating mode tests were accomplished with reversed HTF flow rates:
the evaporator flow rate of test A is the condenser flow rate of
tests C and D; the condenser flow rate of test A is the evaporator
flow rate of tests C and D. The resulting temperature differences
could then be used for the computer simulation runs. Those results
are reflected in table 2.
1 high temperature cooling
2 low temperature cooling
3 high temperature heating
4 low temperature heating
5 volume flow rate condenser (heat transfer fluid)
6 volume flow rate evaporator {heat transfer fluid)
6
-------
Test A1
Test B1
Test C1
Test D1
°C
°C
°C
°C
(°F)
(°F)
(oF)
<°F)
Condenser Inlet
35.0
27.8
21.1
21.1
(95)
(82)
(70)
(70)
Condenser Outlet
43.2
37.4
32.5
28.1
(110)
(99.3)
(90.5)
(82.6)
Evaporator Inlet
26.7
26.7
8.3
-8.3
(80)
(80)
(47)
(17)
Evaporator Outlet
14.4
13.8
2.7
-11.3
(58)
(56.8)
(36.9)
(11.7)
TABLE 2: Heat Transfer Fluid Temperatures used for the Computer
Simulations
2.2.3 Processing the Simulation Results
For each of the selected operating conditions, all possible
binary refrigerant mixtures were simulated for the chosen pure
substances. Each test condition was evaluated at twenty-one
different mixture compositions for each mixture. This is ac-
complished by changing the composition in steps of five percent
from one pure substance to the other. The results of the simula-
tion runs were then processed in a spread sheet fashion in order to
provide a graphical presentation. Simulation data for R22 is
provided as reference in the graphs. This allows further insight
in the relative performance changes with respect to the different
mixture components and the different mixture compositions.
Additionally, a graphical presentation of the condenser pressures,
discharge temperatures, and suction pressures of the various
mixtures was generated.
3. Selection of Mixture Components
To decide which substances to use in the laboratory tests, the
following criteria had to be considered:
• The compressor discharge pressure should not exceed
2600 kPa (377 psia) under any operation condition
1 see footnotes for table 1 on previous page
7
-------
• The compressor suction pressure should always be
above ambient pressure, i.e. above 100 kPa, in
order to insure that no dirt as air or water vapor
enter the system.
• The compressor discharge temperature should not
exceed 150 °C (302 °F) in order to prevent chemical
breakdown of the lubricant.
• The Q2one Depletion Potential of each substance has
to be zero (i.e. chlorine and bromine free) which
leads to HFC's (if only the methane and ethane
series is chosen).
• The Greenhouse Warming Potential should be sig-
nificantly lower than the values for the currently
used refrigerants.
• Critical point values
• Toxicity
• Flammability
The issue of which lubricants are sufficiently soluble in
chlorine-free refrigerants has not yet been determined. Thus, for
this study, a mineral oil with 33 cSt at 37.7 °C (150 SUS at
100 °F) viscosity was selected and used for all tests. This was
satisfactory for the relatively few hours the compressor was
actually run. Since the performances of all different working
fluids were evaluated with the same lubricant, the impact on the
relative performance is considered minimal.
4. Pure Refrigerants
4.1 General Remarks
Considering pure substances with respect to the criteria given in
section 3, there are some statements that can be made to evaluate
the possible usage of these refrigerants. The pure media are also
compared to pure R22 in this section. It should be noted that R22
is the best of the pure refrigerants with respect to the combined
performance in COP and volumetric capacity for the defined
operating conditions. Refrigerants such as R23, R32, R125, and
R143a have a higher volumetric capacity but a lower COP compared to
R22. For R134a and R152a the opposite is true. The COP of these
fluids is higher, but their volumetric capacity is too low to use
these substances as a drop-in replacement for R22. This also
describes the problem in finding a replacement for R22 since R22
appears to be an almost ideal working fluid for the heat pump
applications, balancing equipment size with operating efficiency.
8
-------
The vapor pressure curves of the named refrigerants are presented
in figure 2. This figure can be used to interpret the performance
of the different refrigerants with respect to their volumetric
capacity and their COP. As a rule of thumb, it can be stated that
the further the refrigerant state line lies to the right (higher
boiling point or lower pressure refrigerant), the higher the
achievable COP. The further the line is to the left (lower boiling
point or higher pressure refrigerant), the higher is the possible
volumetric capacity.
The critical point and normal boiling point values given in
sections 4.2 to 4.7 are obtained from reference [4].
The values for the ODP for the following pure refrigerants is
zero, as required by the selection criteria in section 3.
The values for the GWP are relative to Rll (GWP=1.0).
10000
1000
100
10
R22, R23, R32, R125, R134a, R143a, R152a
-90
-60
-30
30 60 90 120
temperature (°C)
Fig. 2: Vapor Pressure Curves of Selected Refrigerants
[ lg p / (1/T) scale]
9
-------
4.2 R23
As a pure refrigerant R23 can only be considered in low tempera-
ture applications, since otherwise the high pressure criterion
specified in section 3 would easily be violated. Its normal
boiling point is 191.1 K. The critical point lies at 298.98 K and
4820 kPa. Because of these constraints, it can only be used in
heat pump applications as a mixture component. R23 has a high
volumetric capacity, which makes it a potentially valuable mixture
component, since it can be used to increase the capacity of low
capacity -but high efficiency- refrigerants. R2 3 is nonflammable
and the toxicity is considered to be low [5], The GWP is about 21
[6] which may eventually rule out any use of R23 as a working
fluid.
4.3 R32
R32 is flammable [7] as a pure substance. Its normal boiling
point is 221.4 K. The critical point values are 351.56 K and
5830 kPa. The computer calculations showed that the volumetric
capacity of the pure refrigerant is significantly higher than that
of R22. As a mixture, it may be a very valuable component if it is
used together with a nonflammable refrigerant, so that the R32
concentration would be outside the flammability limits (controlled
flammability in a mixture) . The toxicity tests for R32 are not
finished [8] but assumed to be low [7]. The GWP for this substance
is 0.13 [9]. These attributes make this refrigerant a very
attractive candidate for mixtures.
4.4 R125
The normal boiling point of R125 is 224.6 K and the critical
point values are 339.4 K and 3631 kPa. The computer s
show that R125 has about the same volumetric capacity but a lower
COP than R22. In fact, the COP of R125 is lower than that of R32
which makes it an exception to the previously stated "rule of
thumb". This is because the critical point is lower, which
typically causes excessive flash gas in the vapor compression
cycle. As a nonflammable refrigerant [7], R125 is a possible
component for refrigerant mixtures. The toxicity tests for R125
are not finished [8] but assumed to be low [7]. The value for the
GWP is 0.58 [9].
4.S R 3.43a
Again, by looking at the log p - (1/T) diagram (Fig. 2), it can
be concluded that R143a has to be very close in its performance to
R22. The normal boiling point is 225.8 K and the critical point
values are 346.25 K and 3811 kPa. With respect to the GWP, R143a
10
-------
has a relatively high value of 0.74 [9]. Rl43a is flammable [7]
and -relying on preliminary data- is probably at least moderately
toxic. With this in mind, R143a does not appear to be a good
choice as a part of any refrigerant mixture. Due to the lack of
information with respect to the toxicological data at this time,
the researchers chose not to perform laboratory tests at NIST with
this refrigerant until further information is available.
4.6 R134&
This is the most readily available alternative refrigerant of the
ones considered here. ' The toxicity tests for this substance are
not finished, but the limited results indicate a low toxicity [8],
Since R134a is also not flammable [73 and its GWP is 0.26 [9], it
is one of the prime candidates of any future refrigerant mixture in
this study. The critical point values are 374.21 K and 4056 kPa.
The normal boiling point is 247.0 K. It turns out to be the only
one out of the six refrigerants in this study that is nonflammable
and has a normal boiling temperature that is higher than that of
R22. This is an important fact with respect to the mixture
flammability, since any mixture that could exceed R22 in volumetric
capacity and in COP would probably be composed of a higher and a
lower boiling substance (compared to R22).
On the lower boiling temperature side of R22, there are two
refrigerants that are non-flammable: R23 and R125.
4.7 R152a
This refrigerant is flammable in the range of 5.1 to
17.1 volume percentage in air [8] and has a low toxicity [7]. The
critical pressure is 4520 kPa and the critical temperature is
386.44 K. The normal boiling point is 249.0 K. The simulation
runs show that R152a has a slightly higher COP and a significantly
lower volumetric capacity than that of R22. The fact of the
flammability of this refrigerant and the lower capacity makes it
difficult to propose its use as a single component refrigerant.
However, in mixtures together with a nonflammable component, R152a
could become a valuable refrigerant if a well-performing nonflam-
mable mixture can be found. Another reason to consider R152a is
the low GWP of 0.03 [9].
11
-------
5. Results of the simulation study
5.1 General Remarks
This section is intended to give the reader the comprehensive
results of the computer simulations.
The best results in the computer study are achieved with the
R152a/R32 refrigerant mixture. Figures A1 to AS in Appendix A show
the calculated results of COP and volumetric capacity change with
respect to R22 versus the mass fraction of R32.
Both refrigerants, R32 and R152a, are flammable, which may be a
drawback of this zeotropic mixture. It should be noted here that
this mixture consists of the two refrigerants with the lowest GWPs
of the considered substances. Depending on the composition of the
mixture, the GWP can range from 0.03 for pure R152a to 0.13 for
pure R32. Compared to R22 (GWP = 0.34; relative to Rll) this means
a significant decrease in GWP.
The second best performing mixture is R32/R134a. For better
f*AWY"iia v* i crtti 4"Viia I m sal nriA -i ci q/4 ¦% *n 4-Vio c* ama 4- n m e sic
wUiu^'clL XoUil| uilcXX XULluCillwC Xd vSbcti wcU Xi* bdillc X X
-------
the graphs showing the COP or the volumetric capacity versus the
mixture composition are presented in Appendix A. Due to the large
number of graphs generated by this study, the authors refrained
from appending all diagrams and restricted themselves to the best
performing mixtures: R32/R152a and R32/Rl34a.
The computer tests did not consider liquid line to suction line
heat exchange (LLMX). For the initial comparison of the mixtures
to the pure R22, it seemed to be appropriate to start with no LLHX.
It should be emphasized again that none of the mixtures con-
taining R23 would be acceptable because of their high GWP. Even
small amounts of R23 for example 5 %-mass would already create a
refrigerant mixture with a GWP of greater than 1.0. On top of
that, the contribution of the other mixture component would have to
be added. The GWP consideration is mentioned here in order not to
repeat it in the sections 5.2.1 to 5.2.5. For all these mixtures,
the GWP would result in unacceptably high GWP's.
Many of the refrigerant mixtures that were investigated in this
study were not considered to be good alternatives for R22. The
discharge pressure criterion of section 3 was often violated for
one or more of the four operating conditions in the composition
range of interest (i.e. COP improvement and/or volumetric capacity
improvement). Thus, these mixtures were rejected as possible R22
replacements under the here considered operating condition.
However, under different operating conditions (i.e. different
applications) other mixtures might perform better than the two
mixtures that were concluded to be best for the here investigated
application in household heat pumps.
5.2.1 R23/R32
This mixture contains two refrigerants that are both low boilers
compared to R22. This translates to high condenser pressures which
exceed the allowable limits in most calculated cases. For test
conditions A and B, there is not a single composition that does not
violate the pressure criterion set for this study in section 3
(maximum discharge pressure of 2600 kPa (377 psia)).
5.2.2 B23/R125
Although R125 is generally lower in its vapor pressure than R32
(Fig. 2) , this mixture also exceeds the pressure limit set for this
study.
13
-------
5.2.3 R23/R143*
The higher boiling pressures of both components compared to R22
make this mixture unattractive considering the criteria of
section 3. Especially in the cooling mode the discharge pressure
limit is exceeded easily.
5.2.4 R23/R134*
This mixture would eventually be suitable with very low R23
content with respect to its projected COP. The calculations show
a maximum in COP at about 10 %-mass of R23 which exceeds the COP of
R22 for test conditions A,B, and C by up to nine percent in the
case of test B. For test D, the mixture COP is about three percent
less than that of pure R22. The volumetric capacity increases
almost proportionally with the R23 content. The capacity of R22 is
met between 15 %-mass and 30 %-mass of R23 depending on the test
conditions. Considering the fact that the test conditions reflect
normal operating conditions rather than extreme conditions, the
mixture would have to contain 30 %-mass R23 in order to achieve the
same heating capacities as R22 under all operating conditions.
That concentration of R23, however, would result in lower COP's and
too high head pressures. Together with the remarks about the GWP
of R23 in section 5.2, this leads to the conclusion that this
mixture is not the desired working fluid for the considered
operating conditions.
5.2.5 R23/RlS2a
The refrigerant R23 again curtails the usefulness of this mixture
for the investigated heat pump application. For low R23 amounts
(under 20 %-mass R23), a distinct maximum in COP can be observed,
which exceeds the R22 COP in the cases of tests A,B, and C by five
to fifteen percent. In the test case D, the COP of the mixture is
slightly lower than that of the pure refrigerant. The COP maxima
all occur between zero and fifteen %-mass of R23. But in this com-
position range, the volumetric capacity of this mixture does not
meet the requirement to be as high as that of R22. At R23 amounts
that ensure a sufficiently high volumetric capacity of the mixture,
the condenser side pressures, as well as a lower COP'S make this
mixture unsuitable under the chosen restrictions for these
operating conditions.
5.2.6 R32/R125
This mixture forms a positive azeotrope (i.e. the vapor pressure
of the azeotropic composition is higher than either one of the
vapor pressures of the pure components). The investigated
operating conditions would already create high discharge pressures
14
-------
for the two pure refrigerants and thus violate the pressure
criterion of section 3. Since the azeotropic saturation pressures
for this mixture are higher than for both pure components (compared
at the same temperatures), it was concluded that this mixture does
not represent the desired working fluid.
5.2.7 R32/R143a
Although the pressures are generally lower than those of the
R23/R143a mixture, the margin of safety with respect to the
discharge pressures seems to be too small, considering extreme
cooling applications.
Another interesting observation can be made by observing the fact
that no maximum in COP appears for this mixture. The higher the
R32 content, the higher the COP (at least for the four test
conditions as long as the pressure limit is not exceeded). This
phenomenon can be explained using figure 2: the property curves of
the pure substances are close together and therefore the tempera-
ture glide during evaporation or condensation due to the zeotropic
effect is small. This, in turn, means that the positive effect due
to temperature glide is very small. Hence, the change in perfor-
mance due to composition shift is approximately a linear change
between the pure substances. The same observation can be made for
the mixtures of R32/R125, R125/R143a and R134a/R152a,
5.2.8 R32/R134*
This combination generally shows good performance as long as the
amount of R32 is kept under 60 %-mass. The highest COP for this
mixture is calculated at 50 %-mass R32 for the cooling conditions
and at 70 and 80 %-mass R32 for the heating conditions (Fig. Al,
A3, A5 & A7). The values are 10% to 12% higher than for R22 in the
cooling mode and six to eight percent higher in the heating mode.
The heating tests show a shallow maxima, which is explained by the
characteristic double maxima of zeotropic mixtures. The benefit of
this broad high COP in this case is that these high R32 contents
are not necessary to achieve a good performance. At these R32 mass
fractions the condenser pressure would easily exceed the pressure
limits in extreme operating conditions. Even mass fractions of
40 %-mass R32 still show four to seven percent COP improvement in
the heating mode.
The volumetric capacity increases with the content of the high
pressure refrigerant R32 (Fig. A2, A4, A6 & A8). Since R134a has
a lower volumetric capacity and pure R32 a higher volumetric
capacity than R22, it is again of interest at which composition the
volumetric capacity of the mixture is higher than that of R22.
This R32 content is 26 %-mass to 27 %-mass for the cooling tests
and 31 %-mass to 35 %-mass for the heating conditions. This
15
-------
implies that for all four test conditions mixtures with at least
35 %-mass of R32 will meet or exceed the volumetric capacity of
R22. At the same time, the COP of R22 will be exceeded.
The flammability aspect of this mixture should again be em-
phasized. R134a as a nonflammable refrigerant would definitely
lower the flammability of the R32/R134a. It has not yet been
established if these compositions are flammable.
As one of the two predicted best performing zeotropic mixtures,
R32/R134a is tested in the MBHP. The results are presented in
section 8 and Appendix B.
5.2.9 R32/R152a
The highest values for the COP are reached between 40 %-mass and
55 %-mass R32 depending on the operating condition (Fig. Al,
A3,A5,A7). An increase in COP of 16% to 18% in the cooling mode
and seven to eleven percent in the heating mode is calculated.
The same volumetric capacity as for R22 is reached at 37 %-mass
and 38 %-mass R32 for the cooling conditions and 44 %-mass and
48 %-mass R32 for the heating conditions (Fig. A2, A4, A6, A8). In
order to assure the same volumetric capacity and the same COP of
this mixture compared to R22, this mixture should contain about
50 %-mass R32.
In the heating tests (Fig. A5 & Fig. A7) R32/R152a shows the
characteristic double maximum for zeotropic mixtures. Due to the
temperature glide during evaporation and condensation, the COP
increases with the R32 mass fraction. The first maximum is reached
at the point of best glide matching. The further increase in R32
content in the mixture causes too much temperature glide on the
refrigerant side. This overglide causes the COP to drop until a
local minimum is reached. That is the composition of worst
overglide. By going to mixtures with even higher R32 mass
fractions, the temperature glide of the refrigerant drops again,
i.e., the temperature glide matching gets better. This way a
second local maximum can be observed at a point of good glide
matching. With a further reduction in refrigerant temperature
glide, the COP goes further down and approaches the value for the
pure refrigerant. This effect can only be observed if at least one
of the refrigerant temperature glides in the evaporator or the
condenser is bigger than the HTF temperature glide. Table 2 shows
the evaporator temperature glide of the HTF for the heating
conditions to be the smallest of the four test series. Test
condition D (low temperature heating) also shows the lowest
condenser HTF glide of the four test conditions. This test
condition shows the two most distinct double maxima in figure A7.
16
-------
Unfortunately, both mixture components are flammable and
therefore have questionable use as replacement refrigerants for
R22. On the other hand, this is the mixture with the lowest GWP
out of the fifteen considered mixtures.
R32/Rl52a is the second mixture that has been tested in the test
rig. The results are compared to test data of R22 and of the
R32/R134a mixture in the later section 8 of this report. The
graphical presentation of the test results can be found in
Appendix B.
5.2.10 R125/R143&
Referring to the remarks about the R32/R143a mixture, the
calculations show that no composition exceeds the highest COP or
the highest volumetric capacity of the two pure substances. At the
same time, both pure refrigerants show a lower COP and a higher
volumetric capacity than R22. Again, the log p - 1/T diagram
(Fig. 2) indicates that the gliding temperature difference for any
mixture composition will be very small. The result is that the
mixture shows performance results between the performance of the
pure substances. But since neither of the pure refrigerants
exceeds the R22 performance, the mixture also does not exceed the
R22 results.
5.2.11 R125/R134a
The performance of this mixture generally does not exceed the R22
performance. Its COP comes close to or even meets that of R22
depending on the composition and test condition. At the same time,
it is, however, not delivering the required volumetric capacity and
therefore does not appear to be a good candidate to replace R22.
5.2.12 R125/R1S2&
This mixture shows good results with respect to the COP, since it
surpasses that of R22 for tests A,B, and C and is almost equal for
test D (in a certain composition range) . But the volumetric
capacity is lower than that of R22 for R125 contents of less than
80 %-mass. In that range though, the COP is decreasing with
increasing R125 amount. The calculated results do not indicate a
mixture composition where both the COP and the volumetric capacity
are acceptable as a replacement for R22.
17
-------
5.2.13 R143a/R134a
The COP of this mixture is about equal to that of R22 for a
50 %-mass to 60 %-mass R134a content. However, in this composition
region of interest, the volumetric capacity is about 30% lower than
that of R22. Therefore, this mixture is not considered a suitable
working fluid for heat pump units.
5.2.14 R143a/R152a
Again, the fact that only one of the two major criteria is
satisfactory curtails the usefulness of this mixture. The COP is
calculated greater than that of R22 for a variety of compositions,
but the volumetric capacity of the mixtures is too small to be
considered as a replacement for R22.
5.2.15 R134a/R152a
The property lines of the two pure components that form this
mixture are again very close (Fig. 2) . That translates to low
temperature glides during the evaporation and condensation. Since
those glides are the basic reason for a performance maximum of a
zeotropic mixture, this characteristic maximum cannot be seen here.
The graphs show almost straight lines for the change in COP and
volumetric capacity with respect to composition change. Since the
volumetric capacity of both single components is about 40% lower
than that of R22, any composition of these two does not deviate
much from that capacity. Hence, this mixture is not considered
suitable for the desired purpose.
6. Discussion of the Simulation Results
The results of this computer study are consistent with the
expectations suggested by the refrigerant component properties.
There are two binary mixtures that indicate a better performance
than the currently used R22. By name, these are R32/R152a and
R32/R134a. Although the pure components of these mixtures show
either lower COPs and higher volumetric capacities or higher COPs
and lower capacities, wide ranges of mixture compositions perform
better in COP and show a higher volumetric capacity.
The mixture of R32/R152a (Section 5.2.9) is definitely flammable
since both pure substances are flammable. Both pure refrigerants
appear to have low toxicities. The GWP of this mixture is the
lowest of all possible binary combinations. The engineering
aspects discussed in section 3 are found to be within acceptable
limits for the investigated composition range.
The calculations for R32/R134a (Section 5.2.7) show a smaller
performance improvement than for R32/R152a but still a significant
18
-------
increase compared to R22. Both of these refrigerants are in the
class of low toxicity and only R32 is flammable. If flammability
tests show that a nonflammable refrigerant mixture exists within
the composition range of performance improvement (compared to R22),
this mixture would clearly be the substance of choice with respect
to safety considerations. The GWP of R32/R134a is higher than that
of R32/Rl52a. However, it is still significantly lower than that
of R22. Again the engineering aspects from section 3 do not appear
to pose any problems.
Given the results of the computer study, the researchers chose to
conduct tests with these two refrigerant mixtures within certain
composition ranges. The ranges of the tested mixtures are largely
determined by the attempt to find a mixture that achieves at least
the same volumetric capacity and the same COP as R22. This should
be the case for all operating conditions. At the same time,
excessive amounts of R32 should not be used in order to provide for
acceptable discharge pressures even under extreme operating
conditions. The researchers chose to run tests for both refrige-
rant mixtures in a range between 15 %-mass and 40 %-mass R32. For
all these compositions, the pressures are expected to be well
within the acceptable range.
19
-------
7. The Mini-Breadboard Heat Pump
7.1 General Remarks
The term Mini-Breadboard Heat Pump (MBHP) originates from the
fact that the charge for the test system is about 0.5 kg (1.1 lb)
of refrigerant. A small required refrigerant charge was a major
design criterion of the test rig since the desired refrigerants are
expensive and/or not readily available.
The test apparatus is designed for a 3.5 kw (1 ton) refrigeration
capacity,
Other design criteria were:
- counterflow heat exchangers (condenser, evaporator and
LLHX)
- liquid heat transfer fluid
variable speed compressor
system usable with and without LLHX
- accessibility of refrigerant and heat transfer fluid in
order to create condenser and evaporator temperature
profiles
7.2 system Design
The system emulates the basic refrigeration cycle. The refrige-
rant side consists of the compressor, the condenser, the liquid
line heat exchanger, the expansion device and the evaporator, as
can be seen in figure 3. The system does not include an oil
separator or an accumulator in order to satisfy the low charge
design criteria.
The refrigerant is compressed in a two cylinder reciprocating
compressor and discharged into the condenser. The condenser is a
counter flow heat exchanger; the refrigerant flows in the inner tube
and a water-ethylene glycol mixture is pumped through the annulus.
In order to increase the heat transfer coefficient, the refrigerant
and the HTF side are modified. The inner tube is equipped with
turbulators1 and the heat transfer on the water-glycol side is
enhanced by using a spine-fin tube. This ensures a turbulent flow
1 twisted tape that was inserted in the inner tube before assembly of the heat
exchanger units
20
-------
on both sides and a large heat transfer area. The resulting
increase in flow pressure drops, especially on the water/glycol
side, was not considered as important as the compactness of the
system design since that allowed for a small refrigerant charge.
After leaving the condenser, the refrigerant flows either through
the LLHX and then to the expansion device or directly to the
expansion device. The liquid line heat exchanger is used to
further subcool the condensed liquid refrigerant. The subcooling
is accomplished with the low pressure refrigerant vapor leaving the
evaporator. The LLHX, as well as the evaporator and the condenser,
is working in pure counterflow.
Cooling
Brine
Charging
System
Three-way
valve
Thermostatic
valve, electric
Thermostatic
valve
Hand valve
vW
vW
HC
HV
Filer, dryer
{refrigerant)
C - condenser
Co -compressor
E -evaporator
FM - volume flow meter
MM - mass flow meter
Fig.
@ Sight glass
| niter {water)
3; System Setup of the Mini-Breadboard Heat Pump
HC - heater, constant
HV - heater, variable
I - Inverter
IX - Internal heat exchanger
M -motor
P -pump
S - torque & speed meter
W/G - water / glycol tank
X - external heat exchanger
The refrigerant enters the evaporator after its expansion in a
manually controlled expansion device. The evaporator is a
duplication of the condenser - the refrigerant again flows in the
inner tube and the water/ethylene glycol mixture in the annulus.
After leaving the evaporator, the refrigerant can be superheated in
the LLHX, as explained before. The refrigeration loop is closed
when the refrigerant vapor enters the compressor again.
21
-------
The heat transfer fluid is, in both cases, a mixture of 60 %-wt
water and 40 %-wt ethylene glycol. This mixture ensures a frost-
proof temperature down to -20°C (-4°F), which makes it possible to
run low outdoor temperature heating tests (Test D in Table 1) . The
same HTF is used in the condenser in order to have similar (i.e.
better comparable) fluid properties in both heat exchangers. This
is desirable because the heating and cooling conditions are
simulated by changing the mass flow rates and operating tempera-
tures of the HTF and not by reversing the refrigerant flow.
The oil problem mentioned in section 3. was more or less ignored.
The system design provided for a good oil return to the compressor
and there have been no problems encountered during the test period.
7.3 System Instrumentation
A data acguisition system is used to measure the engineering data
that are needed for a detailed analyses of the tests. This system
is also used to control the machine during its operation. All
parts of the MBHP such as compressor, heaters, and pumps can be
switched interactively via the computer in connection with the data
acguisition.
During the test period, data are read every minute and can be
compared to the desired set points of the heat transfer fluids.
For a successful test, the data for the set points have to be
constant within reasonable deviations for at least 30 minutes. The
data are then averaged over the time period of the test and used
for the performance evaluation.
The collected data consist of 97 temperatures, 18 voltages, and
2 frequencies. Figure 4 shows the location of the thermocouples,
pressure transducers, flow meters, and all other measuring devices
used to determine the performance of the different refrigerants.
The temperature measurements are conducted with type T ther-
mocouples (copper-constantan).
Appendix C contains an uncertainty analysis for the COP and mass
fraction measurements.
22
-------
—(XH-
Cooling
Brine
Changing
System
©'^H@Y0
Measurements:
(TP) - thermopile
(t) - temperature
(DP) - pressure difference
¦ pressure
S - torque & speed meter
FM - volume flow meter
MM - mass flow meter
C - condenser
Co -compressor
E - evaporator
HC - heater, constant
HV - healer, variable P - pump
I - Inverter W/G - water / glycol tank
IX - Internal heat exchanger X - external heat exchanger
M -motor
Fig. 4: Instrumentation of the Mini-Breadboard Heat Pump
7.4 Test Criteria
The criteria for the tests of the different refrigerant mixtures
are derived from section 2.2.2 and the approach of testing the
substances at a constant — value1 as suggested bv McLinden
and Radermacher [10].
McLinden and Radermacher concluded that an appropriate method of
comparing pure refrigerants with zeotropic refrigerant mixtures is
1 U j heat transfer coefficient (W/(m5*K))
total heat transfer area (ra2)
Q : capacity (W)
23
-------
U*A
to specify the HTF temperatures and to keep the ratio of
Q
constant. This approach is slightly modified in the study in order
to make it more practical. Using the calculated HTF temperatures
in table 2 and R22 as working fluid in the MBHP, the values for the
capacities are experimentally established. For each test con-
dition, the capacity is measured with R22 and held constant for all
mixture tests of the same operating condition. The heat exchanger
areas are also held constant throughout the test series. This
implies that the compressor speed has to be changed from test to
test. A drawback of this method is that the values of the heat
transfer coefficients, U, are not necessarily constant for all
tests. Since the prediction of the heat transfer coefficients for
the different working fluids would be very difficult to accomplish,
the authors decided to accept this deviation from the proposed test
criterion. McLinden and Radermacher mentioned in their paper that
such concessions may have to be made for the actual testing of
mixtures.
In order to get comparable test results, the tests have to be
further restricted with respect to condenser subcooling and
evaporator superheat. The researchers chose to run the tests with
as little subcooling as possible (effectively not greater than 2 K)
and that the superheating region should start at 80% (± five
percent) of the evaporator length. The superheat criterion is
necessary in order to ensure that no two-phase refrigerant is
leaving the evaporator, since that would affect the performance
rating of the system. The subcooling and superheat criteria are
held constant for all operating conditions. Both criteria require
the adjustment of the refrigerant charge during the test procedure.
The charging system design (Fig. 3) is crucial in accomplishing
that task. The capacity defining tests are conducted at compressor
Speeds of 1500 RPM to 1620 RPM.
For the composition measurements, a small vapor sample is taken
from the discharge line during the steady state phase of the test
and then evaluated with the gas chromatograph.
24
-------
8. Test Results
In this section, the test results which are graphically presented
in Appendix B are discussed. The compressor speeds, CQPs, suction
temperatures, discharge temperatures, suction pressures, and
discharge pressures are reported versus the mass fraction of R32,
The suitable R22 data are displayed as a reference line in those
graphs.
8.1 General Remarks
Again, it should be noted that the tests are obtained under the
criterion of constant capacity for all tests in one test series.
This has the effect that different compressor speeds {Fig. Bl.1 to
B1.5) are used for different refrigerant compositions since the
volumetric capacity of a mixture varies with its composition (see
simulation results in Appendix A) . Different compressor speeds, as
well as different refrigerant densities and transport properties,
affect the compressor performance and the pressure drop within the
heat exchangers. But, these changes are system specific rather
than refrigerant specific. Considering a proper system layout for
each refrigerant mixture, it should be possible to reach the same
compressor efficiency, as well as the same pressure drops indepen-
dent of the working fluid. This suggests that a correction of the
test data with respect to the named non-idealities would give some
more insight in the differences of pure and mixed refrigerants.
However, for this study, these thoughts are not reflected in the
results since the process of how to correct the test data has yet
to be determined.
8.2 High Temperature cooling Test (test condition 1A)
The cooling capacity, which is kept constant for these tests, is
3,510 W (11,976 BTU/h). The compressor speed is shown in
figure Bl.l. As one point of interest, the refrigerant mixture
R32/Rl34a containing 24 %-mass R32 should be pointed out because
for that composition the speed and capacity are equal to that of
R22. For the mixture of R32/R152a, that is true at about 31 %-mass
Of R32.
The COP of the R32/R134a mixture (Fig. B2.1 & B3.1) shows a
steady increase in the measured composition range with increasing
R32 amount. For R32 mass fraction below 0.22 in the R32/R134a
mixture, the COP is lower than that of R22. The maximum is
reached at an improvement of 19% over R22 at 40 %-mass of R32. For
R32/R152a, the COP is higher than that of R32/R134a for mass
fractions lower than about 0.37. The tests indicate that the COP
levels out for this mixture at that composition. The maximum
measured increase in COP is 17% to 18% over that of R22 at the same
25
-------
capacity. At the composition that shows the same speed audi
capacity as R22, the COP is 14% higher for R32/R152a and four
percent higher for R32/Ri34a.
The suction pressures (Fig. E4.1) are proportional to the R32
content of the mixture. They vary in the tested range ±20%
compared to R22. The lower pressures at lower R32 content also
mean low volumetric capacities (i.e. higher compressor speed, since
the tests were conducted at a constant capacity). The same suction
pressure as for R22 is reached at about 28 %-mass R32. The R152a
mixture generally shows lower pressures than R134a at the same R32
content. Its mixture pressures are lower by about 10% (with
respect to R22) . At a mass fraction of 36% the suction pressure of
R152a is equal to that of R22.
For the discharge pressure (Fig. B5.1), a similar statement as
for the suction pressure can be made, although the differences are
not as big on a relative basis. For R32/R134a, the discharge
pressure varies from -12% to +8% (with respect to R22) in the mass
fraction range of 0.17 to 0.4. For the R32/R152a binary, the
discharge pressure varies from -24% to -6% in the measured range.
The figures for the suction and discharge pressures show that
both refrigerant mixtures lie well within acceptable pressure
limits (section 3), which was anticipated for this test condition.
With respect to the suction and discharge temperatures (Fig. B6.1
& B 7,1) observed in this test series, it can be noticed that the
suction temperature is constant (within l K) for all tests
independent of the working fluid. The discharge temperature,
however, is not constant. For both mixtures, the discharge
temperature is lower than that of the comparable R22. The
R32/R152a mixture shows a decrease of 3 K to 5 K (with increasing
R32 content) and the R32/R134a mixture shows a decrease of 5 K to
almost 10 K compared to R22. The difference in the discharge
temperatures can be explained with the different specific heats of
the working fluids.
8.3 Low Temperature Cooling Test (test condition IB)
The cooling capacity at which the test results for this test
condition are obtained is 3,685 W (12,574 Btu/h). The obtained
change in compressor speed is shown in figure B1.2. The mixture
composition that is run at the same speed as R22 contains 23 %-mass
of R32 for the R32/R134a mixture and 29 %-mass R32 for the
R32/R152a test series.
The COP of all tested refrigerant mixtures is higher compared to
test condition 1A since the temperature lift is smaller. The
relative and absolute change of COP with respect to R22 is
presented in figure B2.2 & B3.2. The mixture of R32/R152a shows a
26
-------
measured improvement of over 20% for R32 contents higher than
35 %-mass. At about 17 %-mass the COP is equal to that of R22, and
below that the COP is also lower. The mixture of R32/R134a shows
a lower COP compared to R32/R152a in the range of up to 30 %-mass
of R32. Above a mass fraction of 0,35 R32 the COP appears to be
higher than that of R32/R152a. The largest increase in COP for
this mixture is about 22% at 38 %-mass R32.
The suction and discharge pressures (Fig. B4.2, B5.2) are again
well within acceptable limits. With respect to R22, the suction
pressures of the R32/R134a mixture are 10% to 15% higher than those
of the R32/R152a mixture (Fig. B5.2). In absolute numbers, this
means a higher pressure of about 100 kPa (14.5 psi) on the suction
side (Fig. 4,2). However, this does not translate into the same
difference on the discharge side. Here, figure B5.2 shows an
absolute pressure difference of about 200 kPa (29 psi). At the
same time, the discharge pressures for R32/R134a are lower than
those of R22 for R32 contents of less than 30 %-mass. For
R32/R152a, all measured discharge pressures are below the obtained
R22 value.
The suction and discharge temperatures (Fig. 6.2 & 7.2) show a
similar results as for test condition 1A. The suction temperature
is within + 1 K (except one data point). The discharge tempera-
tures are generally lower than the corresponding R22 temperature.
For the R152a mixture, the difference in discharge temperature
varies from 3 K lower at about 20 %-mass to 5 K lower at R32 mass
percentages higher than 30. As for the R32/R134a mixture, the
decrease compared to the R22 discharge temperature is higher and
amounts to about 10 K for mixtures with more than 30 %-mass R32.
8.4 High Temperature Heating Test (test condition 1C)
For this test condition the reference heating capacity is 3,140 W
(10,714 Btu/h). Figure B1.3 shows the compressor speed variation
versus the composition change of the mixtures. For R32/R134a, the
speed equivalence with the R22 test is obtained at 28 %-mass R32,
and for the R32/R152a mixtures this point is reached at 37 %-mass
R32.
In the heating mode, the heating COPs of the R32/R134a mixture
are higher than those of the R32/R152a mixture (Fig. B2.3 & B3.3).
This observation is different from the cooling mode where the
cooling COPs of R32/R134a are lower compared to R32/R152a. The
increase in COP of R32/R134a is about two percent based on the R22
COP. For R32 contents of more than 20 %-mass, the COP is higher
than that of R22 for the R134a binary. The R32/R152a mixture also
shows an COP improvement against R22 above 20 %-mass R32. At the
mixture compositions of same speed and same capacity, however,
R32/R134a shows four percent improvement over R22 and R32/R152a
shows eight percent.
27
-------
The suction and discharge pressures at this operating condition
appear lower than the R22 tests for less than 32 %-mass in the
R32/R134a mixture (Fig. B4.3, B5.3). The R32/R152a mixture again
shows generally lower pressures and never reaches the R22 discharge
pressure in the tested range.
The suction temperatures in figure B6.3 appear to be slightly
lower (about 1.5 K) than in the R22 reference test. However, for
the two mixtures, the suction temperature is almost constant over
the complete tested composition range.
The discharge temperatures {Fig. B7.3) prove to be about 5 K
higher for the R32/R152a mixture than the R32/R134a binary. For
both test series the discharge "temperature is lower than the
reference temperature. The difference is from 3 to 5 K for the
R152a binary. For the R134a mixture tests, the temperatures are 10
to 8 K lower for the test over 2 3 %-mass R32.
8.5 Low Temperature Heating Test (test condition ID)
These tests are conducted at a condenser capacity of 1,970 W
(6,722 Btu/h). Xn figure B1.4, the speed variation is shown with
respect to the measured compositions of the mixtures. For
R32/R134a the point of same speed and same capacity can be
identified at about 35 %-mass R32. The R152a binary fulfills this
criterion at 45 %-mass R32.
Similar to the high temperature heating tests, it appears that
the R32/R134a mixture performs better in the heating mode than the
other binary mixture at the same R32 content. With respect to R22,
the R134a binary is about five to ten percent better than the R152a
binary (Fig. B2.4 & B3.4). The same COP as for R22 is reached at
about 30 %-mass R32 for the R32/Rl34a mixture. The Rl52a binary
needs about 40 %-mass R32 to reach the COP of R22. In the tested
composition range, only R32/R134a shows an improvement in COP of
five percent over R22 at 40 %-mass R32. At the composition of
equal speed and capacity with the R22 reference test, both mixtures
show the same COP as R22 for this test condition.
The measured pressures in figure B4.4 and B5.4 again show
R32/R152a to be the low pressure mixture (compared at the same R32
content). It never reaches the same suction or discharge pressure
as R22. The R32/R134a mixture pressures are relatively close to
the R22 pressures and range between ±10% for the discharge pressure
in the measured composition range. The suction pressures are lower
than the ones for R22 in the range of less than 30 %-mass R32.
For the suction and discharge temperatures shown in figure B6.4
and B7.4, a slightly different behavior than before can be
observed. The discharge temperatures exceed the reference
temperature of R22 for certain mixture compositions. The discharge
28
-------
temperature of the R152a binary generally exceeds the R22 tempera-
ture by 13 K to 3 K in the range of 20 %-mass to 43 %-aiass R32.
For R32/R134a, a higher discharge temperature than for the R22
reference test is measured for R32 mass fraction of less than 0.23.
For mass fraction above 0.23 R32, a lower discharge temperature is
measured (5 K to 8 K) . The suction temperatures are again
relatively constant for each mixture series (within 2 K) but the
Rl34a mixture tests show a lower temperature (2 K to 3 K below R22
reference and 1 K to 2 K below the R152a series).
8.6 Liquid Line Heat Exchange (test condition 1A-LLHX)
This test series has been conducted to find eventual benefits
associated with the use of the LLHX. The tests are taken at the
same operating conditions as test series 1A (high temperature
cooling). This also means that the same evaporator capacity of
3,520 W (12,011 Btu/h) is used. The reference values for R22
referring to this section are also obtained with a LLHX.
Figure B1.5 shows the compressor speed variation for both
mixtures. The point of equal speed and capacity is reached at
24 %-mass R32 for the R32/R134a mixture and at 32 %-mass R32 for
the R152a binary.
The COP in figures B2.5 & B3.5 increases for both mixtures with
the addition of R32. The tests show an almost linear increase of
COP for the R134a mixture matching the COP of R22 at 20 %-mass R32.
At 40 %-mass R32, this mixture shows a 17% higher COP than R22.
For R32/R152a, the slope of increase in COP appears to be smaller.
Although the COP is higher for R32, mass fractions below 0.3, a
lower COP than for R32/Rl34a can be observed for mass fractions
over 0.35 R32. The absolute values for the COP reach 4.4 for
R32/R152a and 4.55 for R32/R134a. For the test without LLHX, the
highest COP values for this operating condition are 4.23 for
R32/R152a and 4.3 for R32/R134a. The results are, in both cases
(R32/R134a & R32/R152a), obtained for capacities of 3520 W (±1 %).
Compared to the tests without the LLHX, a significant performance
improvement was measured using the LLHX (+4% in COP for R32/R152a
and +5.8% in COP for R32/R134a) . However, it should be noted again
that part of the performance improvement with the LLHX has to be
attributed to the lower compressor speeds that are needed to run at
the same capacity. By using the LLHX, the latent heat available
during evaporation increases, thus requiring a lower refrigerant
mass flow rate if the capacity is held constant. Although the
suction gas density drops due to the superheat created by using the
LLHX, a lower compressor speed was measured for tests using the
LLHX (comparing tests with the same refrigerant composition at the
same capacity) (see figure Bl.l & B1.5). The latent heat effect
also appears to be the reason why the COP of the R22 tests with
LLHX increases compared to the R22 tests without LLHX. In order to
29
-------
give an accurate assessment of the influence of the LLHX, it will
be necessary to analyze the obtained data further.
At the composition of equal speed and capacity with R22, the COP
improvement of the R32/Rl34a mixture is three percent and of the
R32/R152a mixture is eleven percent.
One interesting aspect of the liquid line heat exchange should be
noted here. When the LLHX is incorporated within the system, the
performance advantage of the R32/R152a mixture over the R32/R134a
mixture in the composition region below 35 %-mass R32 (Fig, B2.1)
shrinks to almost zero. In fact, it seems that in the range above
30 %-mass R32, the mixture of R32/R134a performs better in the
cooling mode. If the phenomenon of a higher increase in perfor-
mance for the R134a mixtures with the usage of a LLHX is true in
the heating mode, then R32/R134a would probably be the refrigerant
mixture of choice.
The suction pressures (Fig. B4.5) for the R134a mixtures are
lower than the ones for R22 in the composition range below
17 %-mass R32. Above that composition, they exceed the reference
pressure by up to 20% at the highest measured composition of
40 %-mass R32. The R152a mixtures are again generally lower than
the ones of R134a and reach the suction pressure of R22 at about
38 %-mass R32. Below that composition, the pressure falls up to
20% below the value of R22. The use of the LLHX increases the
absolute suction pressures for this test series which is expected
from the theory of the liquid line heat exchange with zeotropic
mixtures.
The discharge pressures (Fig. 5.5) are almost identical to the
ones measured for the tests without the LLHX. R32/R134a reaches
the R22 pressure at 29 %-mass R32. R32/R152a is about 300 kPa
(43.5 psi) lower in its condenser pressures at the same R32
content.
Compared to the tests without LLHX (Fig. B6.1) the suction
temperatures are about 10 K higher for the tests with LLHX
(Fig. B6.5). The start of the superheat region in the evaporator
is kept the same as for the tests without the LLHX. Comparing the
temperatures to the R22 reference test, both mixture series show a
2 K to 3 K lower suction temperature.
The discharge temperatures in figure B7.5 are naturally also
higher with the use of the LLHX. The R134a mixtures show the
lowest temperature difference with about 10 K less than R22. The
R152a test shows about 5 K less than R22. All discharge tempera-
tures are well within acceptable limits.
30
-------
8.7 Test Conclusions
From the graphical presentation of the results, it can be
concluded that both refrigerant mixtures are suitable replacements
for R22 if the appropriate mixture composition is chosen and
counterflow heat exchange with the heat transfer fluid is possible.
For R32/R152a, the mixture should have a R32 content of at least
40 %-mass R32. This high fraction of R32 is necessary to ensure a
heating capacity that meets the R22 capacity for all operating
conditions (limiting is test condition ID). At all other operating
conditions, a significant increase in COP can be expected given
counterflow heat exchange. The operating pressures and tempera-
tures indicate no problem in the usage of this mixture. This
mixture has the lowest GWP possible of the tested mixtures which is
about one-fourth of the R22 value. However, this zeotropic mixture
is flammable in the whole composition range.
The other zeotropic mixture investigated has been found to
perform not as well as R32/R152a in the region of lower R32 content
if used without LLHX. But, for mixtures consisting of at least 35
%-mass of R32, this R32/R134a zeotrope shows better or equivalent
performance than R32/R152a. There are two definite advantages of
this mixture. The performance in the heating mode is better than
that of R32/Rl52a since, in general, this mixture has a higher
volumetric capacity at the same R32 concentration. This is
important since it affects the need for supplementary heat
(resistance heating) during the heating period. The second reason,
and this might be even more important, is the flammability aspect.
Since R134a is not flammable, this is a 'controllable-flammable'
mixture, meaning that a certain range exists in which the mixture
is nonflammable. Flammability tests with this mixture are being
conducted at the current time at NIST. Interesting to note is that
the computer simulations indicated a better performance for the
R32/R152a mixture over the whole composition range. This was not
found to be the case for the test results. This deviation from the
computer prediction can probably be attributed to the differences
in operating parameters as discussed in section 2.
The tests with LLHX favor the mixture of R32/R134a since,
compared to R32/R152a, they show a higher increase in the COP with
respect to the tests without the LLHX. The possible benefits of
the liquid-line heat exchange are remarkable with respect to
another aspect. These performance increases are not dependent on
the kind of heat exchange used in the evaporator or condenser. As
a separate counterflow unit in the refrigeration cycle, the LLHX
impact on performance is independent of the kind of evaporator or
condenser used. If these increases in COP with the usage of the
LLHX can be validated for other test conditions, then there are two
new aspects. One is that with the same R32 content higher COPs can
be achieved. The other aspect is that, if the flammability is an
important issue, the R32 content can be lowered to a point where
31
-------
the volumetric capacity is still satisfactory arid the mixture is
nonflammable. For this decision, however, it is necessary to know
the flammability region of the mixture.
The increases in COP over R22 that were measured with both
mixtures amount to up to 20% (test condition B; about 40 %-mass
R32, no LLHX), With cross or parallel flow heat exchangers and
with a design resulting in the same pressure drops than for R22,
these high increases can not be expected. However, the significant
improvements that were measured offer enough potential that even
cross or parallel flow heat exchange (as used in household heat
pump units) should benefit from the usage of these mixtures.
Considering the use of a LLHX for the mixture of R32/Rl34a, it is
very likely that a significant increase in COP remains even with
heat exchangers that do not use the temperature glide of the
mixtures (i.e., cross flow and parallel flow heat exchangers).
This is the case, since a LLHX can always be incorporated in a
system in counterflow. The benefits that are due to the internal
heat exchange persist independently of the kind of evaporator or
condenser that is used. The advantage due to the implementation of
the LLHX in a system is estimated to be about five percent for the
R32/Rl34a mixture (comparing at the same mixture composition with
and without LLHX).
The test results for all mixtures and compositions were obtained
using the same test apparatus. There was no optimization of the
test equipment with respect to pressure drop, compressor efficien-
cy, heat exchanger surface area, etc. for any specific working
fluid. This is important because it turns out that the pressure
drop in the heat exchangers for the R32/R134a mixture is sig-
nificantly higher than that of the R32/R152a mixture (compared at
the same R32 mass fraction) . This pressure drop is system
dependent, not refrigerant dependent. The compressor, however,
does not differentiate operating pressures that are created due to
pressure drops or due to the fluid properties. Thus, the pressure
difference that the compressor has to overcome is increased due to
the pressure drops in condenser and evaporator. These pressure
drops were significantly higher for the R32/R134a mixtures than for
the R32/R152a mixtures {at the same R32 content) thus creating a
disadvantage for the R32/R134a mixtures. Exactly how big a
disadvantage has not yet been determined. For a real system the
pressure drop is a design criterion, therefore, for both mixtures
the same pressure drops can be expected.
9. Comparability of Computer study and Test Results
The chosen path of using computer simulations to find a well-
performing refrigerant mixture to suit heat pumping needs turned
out to be efficient and accurate within the expected limits of the
computer model. The two chosen mixtures were predicted to perform
32
-------
better than R22 assuming the same speed, pressure drops, and
compressor efficiencies. In general, this performance increase has
been found to be true. If the predicted relative changes are
compared with the relative changes of test results taken at the
same compressor speed (the R22 test speed), it can be found that
the simulated improvements in COP are very close to the ones
measured. This further establishes the confidence in the NIST
simulation model.
If the predicted relative changes are compared with the measured
relative changes at compressor speeds deviating from the R22 test
speed, large discrepancies can be noticed (see also section 2).
These discrepancies can be explained by changing compressor
efficiencies and significantly different pressure drops for the
different tests. For higher compressor speeds (i.e., lower
volumetric capacity mixtures), the average fluid velocity in the
system is higher. That causes higher pressure drops for the same
operating condition. As a result, the compressor has to pump the
refrigerant over a larger pressure difference: i.e., for the same
tests at constant capacity the COP decreases. The opposite
scenario is true if mixtures with higher volumetric capacities are
used.
It should be clear to the reader that neither a constant speed
nor a constant capacity test will satisfy the restrictions used in
the computer simulation and that it was not intended to be that
way. By choosing either test criterion, the comparability of the
actual test results is limited. This is the case, since in both
tests the pressure drop in the heat exchangers changes with the
changing working fluids due to compressor speed variations and/or
refrigerant density variations. For the constant speed test, a
change in capacity (i.e., in the heat exchanger effective tempera-
ture difference and/or in the heat transfer coefficient) further
misrepresents the performance changes in the refrigeration system
due to the different working fluids. However, if the test data are
corrected with respect to changes in compressor efficiency and
system pressure drops, the comparability can be regained. Since
the simulation does not account for transport properties and the
test apparatus is not redesigned for each mixture, such differences
have to be expected. This constitutes the difficulty with the
chosen approach, since the differences of the theoretical input
data and the actually measured data are the reason for a substan-
tial deviation in the performance of the considered refrigerant
mixtures. This is specifically true for the pressure drops in the
heat exchangers which are assumed to be constant for the simulation
runs. However, knowing those limitations, the results are still
very powerful and provide the basis for the conclusions drawn in
this analysis.
33
-------
10. Conclusions
This study shows that the two zeotropic refrigerant mixtures,
R32/R134a and R32/R152a, may be considered as replacements for R22
if the appropriate mixture compositions are chosen. Data indicate
that multiple tradeoffs exist in mixture performance for different
system compressor speed and mixture compositions.
If results for R32/R152a are compared at the same compressor speed
and capacity at which the R22 results were obtained, the predicted
improvements in COP are very close to the measured COP's. These
improvements over R22 range from 14 percent for the high tempera-
ture cooling mode to two percent for low temperature heating mode.
Operating pressures and temperatures of this mixture are well
within acceptable limits. The Global Warming Potential of the
tested mixture is about one-fourth the value of R22. However, this
zeotropic mixture is flammable in the whole composition range.
The other mixture, R32/R134a, is not predicted to perform as well
as the R32/Rl52a mixture. However, the test results indicate, that
for mixtures containing more than 35 mass percent of R32, the
performance is very similar to that of the R32/R152a mixture. In
the heating mode, slightly higher COP's were obtained and in the
cooling mode, slightly lower COP's were measured (compared to the
R32/Rl52a mixture at the same R32 mass fractions). The performance
improvement over that of R22 at the speed and capacity equivalent
to the R22 tests ranges from five percent in the high temperature
mode to two percent in the low temperature heating mode. Since
R134a is not flammable, the R32/R134a mixture exists in a certain
range as a non-flammable mixture. Flammabiiity tests will be
conducted to determine this range. The test results for the
R32/R134a mixture show no problems with respect to extreme
pressures or temperatures in the tested composition range.
The achieved COP increases for both mixtures offer enough potential
that even cross flow heat exchange systems (as currently used in
household heat pumps) should benefit from the usage of these
mixtures. This is espacially true for the R32/R134a mixture
considering the use of a LLHX.
It should be stated that the test results of the mixtures were
compared with those of R22 at the same heating/cooling capacity.
The test results for the different working fluids were achieved
with the same test apparatus, meaning there is no optimization with
respect to system pressure drops, compressor efficiency, etc.
34
-------
11. Further Research
As reported in this study, both refrigerant mixtures are composed
of at least one flammable component. For the R32/R134a mixture it
seems to be a logical step to find the flammability region. Those
results will provide the information up to which R32 content the
mixture is nonflammable. If this range includes the 35 %-mass or
even 40 %-mass R32 composition, then this mixture has all the
features needed for a wide acceptance.
In this report the approach of;
— =ATBff=const. (l)
n*A ott
u*nCoCal
is chosen (see section 7.4 for more information on the test
criterion) . In the tests, this is not accomplished in a strict
sense since only Q 1 and A,^ are Jcept constant. The need to define
test conditions for such studies more closely and to describe a way
of interpreting test data is apparent in this study and deserves a
closer look in future worJc.
The changes in compressor speed with changing mixture com-
positions affect the performance of different refrigerant mixtures
due to changing pressure drops in the heat exchangers and due to
changing compressor efficiencies. These effects should be
extracted from the results in further examination of the test data
in order to get a result based on the performance of the different
working fluids and not on their performance in a certain system.
1 U s heat transfer coefficient (W/(m2*K))
A ¦; total heat transfer area (m2)
Q i capacity (W)
ATefr s effective mean temperature difference in the heat exchanger
35
-------
12. References
[1] Domanski P.A., and McLinden M.O. , A Simplified Cycle Simu-
lation Model for the Performance Rating of Refrigerants and
Refrigerant Mixtures, ASHRAE Purdue CFC Conference, 1990, July
17-20
[2] Morrison, G., and McLinden, M.O. , Application of a Hard Sphere
Equation of State to Refrigerants and Refrigerant Mixtures,
Nat. Bur. Stand. (U.S.) Tech. Note 1226, 1986 141 p.
[3] ASHRAE STANDARD ANSI/ASHRAE 116-1983, American Society of
Heating, Refrigerating and Air-Conditioning Engineers, Inc.,
pg. 17, 1983
[4] McLinden M.O., Thermodynamic Properties of CFC Alternatives :
A Survey of Available Data, International Journal of Refrige-
ration, Vol. 13, p.149-162, 1990 May
[5] ASHRAE Thermodynamic Properties of Refrigerants, American
Society of Heating, Refrigerating and Air-Conditioning
Engineers, Inc., p. 170 and p. 191, 1969
[6] Hampson,R., Personal conversation, National Institute of
Standards and Technology, Gaithersburg, MD, 1991 April
[7] McLinden M.O., and Didion D.A., CFC's - Quest for Alternative
Refrigerants - A Molecular Approach, ASHRAE Journal, V.29,
No.12, pp.32 - 42 ,1987 December
[8] Hoffman J.S., Replacing CFCs: The Search for Alternatives,
AMBIO, Vol. 19, No.6-7, 1990 October
[9] Kurylo M.J., Atmospheric Characteristics of HFC-32 (CH2F2) ,
Symposium to Evaluate R-32 and R-32 mixtures in Refrigeration
Applications, EPA, 1991 March 19
[10] McLinden, Mark O., and Radermacher, R., Methods for Comparing
the Performance of Pure and Mixed Refrigerants in the Vapor
Compression Cycle, International Journal of Refrigeration, V.
10, No. 6, pp. 318-325, 1987 November
[11] Doeblin, E. 0., Measurement Systems, revised edition, McGraw-
Hill, p. 38-68, New York, 1975
36
-------
APPENDIX h
SIMULATION DATA
The data presented in this appendix show simulation results
obtained with the program CYCLE11.
37
-------
se
w
CM
(M
a
L.
ai
>
o
+j
c
a>
E
0)
>
0
a
o
u
I
S32/R152a
R32/R134B
/ /
7
0.4 Q. 6
mass fraction R32
Pig. Al:
Relative COP of R32/R134a & R32/R152a vs. R32
content; high temperature cooling conditions
60
r">
&
50
C\J
40
(.y
€L
\
30
Q)
>
0
20
<13
w
ro
10
w
<
c
>.
4-J
- 10
<>
(0
-2D
0
CO
u
- 3D
o
--ID
>
- 5Q
R32/FH3*a ~
s R32/Pi52a
X
x
X //
Fig. A2;
0.4 0,5
mass fraction R32
Relative vol. capacity of R32/R134a & R32/R152a
vs. R32 content; high temperature cooling con-
dition
38
-------
mass fraction R32
A3: Relative COP of R32/R134a & R32/R152a vs. R32
content; low temperature cooling conditions
\
\
R32/R134a^^
^^32/P152e
O 0.2 0.1 a.S 0.8 1
mass fraction R32
A4: Relative vol, capacity of R32/R134a & R32/R152a
vs. R32 content; low temperature cooling con-
dition
39
-------
R32/ R152a
1332/R1 34a
D 4 Q,S
mass "fraction R32
o. B
Fig. A5:
Relative COP of R32/R134a & R32/R152a VS. R32
content; high temperature heating conditions
BO
40
20
R32/R134a^--^
""R32/ R15ZB
Fig. A6:
D,A 0,6
mass Traction R32
Relative vol. capacity of R32/R134a & R32/R152a
vs. R32 content; high temperature heating condi-
tion
40
-------
R32/Rl52o
R32/ R134S
7
0 0.2 0.4 0.6 0.S 1
mass fraction R32
Fig. A7: Relative COP of R32/R134a & R32/R152a vs. R32
content; low temperature heating conditions
^R32/R134a^^
R152a
y
~ 0,2 0.4 D.6 O.B 1
mass fraction R32
Fig. A8: Relative vol. capacity of R32/R134a & R32/R152a
vs. R32 content; low temperature heating condi-
tion
41
-------
APPENDIX B
TEST DATA
The data presented In appendix B is data obtained in test using
the refrigerant mixtures R32/R134a and R32/R152a as well as the
pure refrigerant R22.
42
-------
2200
2100
R32/Rl34a
Fig, Bl. 1:
# R32/Rl52a
R22 reference
110D
0. 15
test condition ia
0,25 0.3 0,35
mass fraction R32
Compressor speed vs. R32 mass fraction;
high temperature cooling condition (1A)
2300
2200
1700
1200
1100
R32/R131a
& R32/R152®
R22 reference
0,1 D.15
test condition iB
0.2
mass
D. 25
tract
0,3
on R32
Fig, Bl.2;
Compressor speed vs. R32 mass fraction; low
temperature cooling condition (IB)
43
-------
2400
/-X
e
a
"O
a>
Q)
a
1/3
L_
o
w
w
Q>
i_
Q.
E
O
U
1800
1600
0 R32/R134a
R32/R152a
R22 reference
0.2 0.25 0.3
mass fraction R32
test condition ic
Fig. B1.3:
Compressor speed vs. R32 mass fraction;
high temperature heating condition (1C)
3000
E 2500
a
W
"O
-------
2000
a
T5
S22
reference
m
i i i t i
0.25 0.3 0,35
mass fraction R32
test conaitlon 1A
Fig. B2.1:
Cooling COP vs. R32 mass fraction; high
temperature cooling condition (1A)
45
-------
5.5
Q.
O 4,5
o
o>
c
o
o
u
R32/R1S2a
I B3a/ni34a
R22 reference
0.1 0.15
lest condition IB
D. 2 0 .25 0.3
mass fraction R32
0,45
Fig. B2.2:
Cooling COP vs. R32 mass fraction; low
temperature cooling condition (IB)
5,3
a
o
o
4 ,4
0*
c
+J 4.2
(d
0)
JZ
4
3.4
0.1 0.15
test conaitTor 1C
B3Z/R134©
$ R32/Rl52a
R22 reference
0,2 D.25 0.3 0.35
mass fraction R32
0,4
Fig. B2.3:
Heating COP vs. R32 mass fraction; high
temperature heating condition (1C)
46
-------
3.4
R22 reTerence
R32/R134a
Q- 3.2
a
U
~fr
C 3
4-J
td
0) , Q
R32/R152a
2.6
2 . 2
015 0.2
test condition
Fig. B2.4:
D. 25 0.3 0.35
mass fraction R32
0,45
Heating COP vs. R32 mass fraction; low
temperature heating condition (ID)
4,6
a 43
O
O
4.2
0>
c
4. 1
0
o
o
3.6
-
0
-
~
E
0 ~ $
R32/R152a
-
m
~
&
B &32/Rt34a
~
R22 reffirence
m
i i i i i
~ .15 0 ¦ 2
test, corvdltion 1A-LLHX
0,25 0-3 0.35
mass fract i on R32
Fig. B2.5:
Cooling COP vs. R32 mass fraction; high
temperature cooling condition with LLHX
(1A-LLHX)
47
-------
r\
w
CM
OJ
cr
0)
>
o
¦M
C
O
L.
a
Q_
O
O -S
-10
~
G!
~ 0
&
£
)
fl32/R152a «
R32/R134&
*
m
¦
m
0.15 0 . 2
test condition 1A
0.25 0.3 0,35
mass fraction R32
Fig. B3.1:
Relative cooling COP vs. R32 mass fraction;
high temperature cooling condition (1A)
CM
r\j
cc
s_
<5!
>
o
c
E
o
u
a
Q.
o
o
10
-zn
0
~
R
0 ~
~
~
B
1
0,1 0,15
test, condition 10
0.3 0.25 0.3 C 35
mass fraction R32
Fig. B3.2:
Relative cooling COP vs. R32 mass fraction;
low temperature cooling condition (IB)
48
-------
10
CM
C\J
CC
>
o
c
CD
E
O
a
a
O
o
b
*
fc3
R3 2/
3134b E
H 4> H32/R15;
«
a
~
m
0.15 0.2 0.25 a.3
mass fraction R32
0. 35
0,45
tiest condition 1C
Fig. B3.3:
Relative heating COP vs. R32 mass fraction;
high temperature heating condition (1C)
r~\
a?
i_j
CM
CM
oc
>
0
4-1
c
01
E
QJ
>
o
L.
a
a
o
o
-10
-25
-aa
R32/R'
a
34a
a
I
!32/R152a
ST
0.1 0,15
test condition 1D
Fig. B3.4:
0,2 0.25 0.3 0.35
mass fraction R32
0.4
0.45
Relative heating COP vs. R32 mass fraction;
low temperature heating condition (ID)
49
-------
Q
w
CM
CM
cr
>
o
•M
C
>
O
i_
a
o_
o
u
15
-10
1!
I
E
S
$
R32/R1
52&
4
4
*
E3 WdtZ/ R1343
s
0,1 ~ 15
t€>st condition 1A-LLHX
~ . 2 ~ , 25 O . 3
mass fraction R32
Fig. B3.5:
Relative cooling COP vs. R32 mass fraction;
high temperature cooling condition with
LLHX (1A-LLHX)
rs
td
a.
v
0)
a
c
o
•M
u
n
w
650
RZ2 reference
350
~ . 15
test, condition 1A
R32/ Rl3-ia
032/0152®
0.25 0.3 0 35
mass fraction R32
Fig. B4.1;
suction pressure vs. R32 mass fraction;
high temperature cooling condition (1A)
50
-------
?aa
R22 reference
R32/R134a
* «32/«152a
0.2 ~. 25 0 3
mass Traction R32
X&&X conoi 1 x 1 on IB
Fig. B4.2: Suction pressure vs. R32 mass fraction; low
temperature cooling condition (IB)
500
r\
fd
CL
w
~
W
W
<1>
L.
Q
U
Z2
II)
350 -
250
R22 reference
P32/R134a
R32/R152*
~ . 1 G- 15
test condition 1C
0,2 0,25 0.3 D.35
mass fraction R32
Fig. B4.3:
Suction pressure vs. R32 mass fraction;
high temperature heating condition (ie)
51
-------
350
ro
a
v
w
Q)
L.
12
tfi
tn
Q)
Q
C
o
-H
U
3
tn
R22 reference
*
#
R32/Rl3«e S3
*
R32/RlS2e
D.25 O.3 D.35
mass fraction R32
test condition 1D
Fig. B4.4:
Suction pressure vs. R3 2 mass fraction; low
temperature heating condition (ID)
R3Z/ R-!3^a
/S
m
S 650
w
®
3 600
0>
ID
0)
Q. 550
C
o
R22 reference
# R32/RT52a
O
13
y>
400
~ . 15 0-2
tsBt confl I t fori 1A-LLHK
0.25 D-3 0-35
mass fraction R32
Fig. B4.5: Suction pressure vs. R32 mass fraction;
high temperature cooling condition with
LLHX (1A-LLHX)
52
-------
2100
m
O-
oj
zs
(SI 17DD
W
Q)
L.
^ 1600
1300
L_
(0
.c
u
{/} 120D
w25> reference
R32/R^34a
*
R32/R152©
0.3 0 . 23 0.3 O . 35
mass fraction R32
Test, condition IB
Fig. B5.2:
Discharge pressure vs. R32 mass fraction
low temperature cooling condition (IB)
53
-------
1500
r\
L.
Z5
w
w
Q)
i_
a
L.
cd
JZ
u
ai
T3
P22 fefer^nce
R32/ni34o
R32/R152a ~
025 0.3 0.35
mass fraction R32
test condition 10
Fig. B5.4:
Discharge pressure vs. R32 mass fraction
low temperature heating condition (ID)
54
-------
Q)
U) 15O0
1200
-
R32/R134C
E
-
R22 ,r«f«fence
a
a sb
-
m
# R32/RtS2a
-
E
~
~
~
~
G **3 O . 2
cond itlon 1A-LLHX
~ .23
mass
0.3 0.33
:t i on R32
Fig. B5.5:
Discharge pressure vs. R32 mass fraction;
high temperature cooling condition with
LLHX {1A-LLHX)
r\ 2B
p
W
-------
30
O
U)
20
R32/Rl52a
"is—"®r"
R22 reference
R32/R134®
_]_
0.1 o. '
test cond11 i on IB
D.25 D.3
mass fraction R32
Fig. B6.2: Suction temperature vs. R32 mass fraction
low temperature cooling condition (IB)
/~\ 13
U
R22 reference
R32/RiS2a <$~
B
R32/ Rl34a
0-1 0 . "13
test, corvd 111 on 1C
O.2 D.23 C.a
mass fracti on R32
Fig. B6.3:
Suction temperature vs. R32 mass fraction
high temperature heating condition (1C)
56
-------
s
r\
o
o.
c:
o
u
D
U)
RS2 reference
R32/R152®
R32/B134a
~ . 25 O . 3 Q.35
mass fraction P32
test condition -ID
Fig. B6.4:
Suction temperature vs. R32 mass fraction;
low temperature heating condition (ID)
O
W
-------
r~\
O
3
4->
o
0)
u
3
¦M
<0
L.
0>
<1)
•M
-------
85
w SO
-------
100
G22 «*-encer
* «32/Rl52e
-------
APPENDIX C
UNCERTAINTY ANALYSIS
C.1 Symbols Used in the Uncertainty Analysis
a Component as R32
A Area
b Component Jb; R134a
cal Calibration Result
COP Coefficient of Performance
cp Specific Heat Capacity
d Total Derivative
e Relative Error
eva Evaporator
con condenser
E Absolute Error
f () Function of .. .
m Mass Flow Rate
mw Molecular weight
N Dependent Variable
13 Compressor Speed
Q capaci ty
T Temperature
T0 Environment Temperature
Tor Compressor Input Torque
V Vol tage
W Compressor Work
x Independent Variable
X Mass Fraction
d Partial Derivative
5 Insulation Thickness
A Absolute Difference
X Conductivity
C.2 General Remarks
The uncertainty analysis reported in this appendix reflects the
effort to provide the reader with information about the accuracy of
the measured data. Since the significant results for this report
consist of COPs, all measured data involved in calculating the
heating or cooling COP are evaluated with respect to their
uncertainty. Quality assurance measures that were taken to confirm
or correct manufacturer data are noted in the sections concerning
those devices. A large number of devices were manufacturer
calibrated and used the first time, so that verification of the
instrument output was chosen rather than a calibration.
61
-------
C.3 Theory
Doeblin [11] presents a procedure for the calculation of the
accuracy of a calculation that depends on several measurements
which are subjected to individual accuracy. The calculated
quantity N is a function of j independent variables x, i.e.,
N = f (xlfx2,x3, . . . ,xA
(I)
The x's are the measurements that are used to calculate N. Each
x is in error by ±dx,, ±dx2, ±dx3,
+dxjf
where the dx values
represent ±3 standard deviations, respectively.
Ej, of N is than given by equation II:
B,
N
The maximum error
(II)
However, the maximum possible error calculated using equation II
does not predict the error very accurately. The root sum square
error which is calculated using equation III is a more realistic
approach to predict the possible errors than using equation II.
With the dx values representing ±3 standard deviations, it can be
assumed that 99.7% of the measured values, N, fall within the error
limits.
-w
df
df
(+ (-^dx2) + (
dx,
df
dx.
dx,)
+
(III)
The absolute errors calculated in equation III can be converted
to relative errors using equation IV. The relative error, eN, has
the units of percent.
*N
En
—Z * 100
N
(IV)
The authors choose to present the uncertainties of values N as:
(v)
N = f {x1,x2,x%,
= f (xltxzrx3,
. ,Xj) ± En
.rxJ ± e
N
C. 4 Heat Transfer Fluid Temperature Difference Measurement
The temperature difference across the heat exchangers and heaters
in the system is determined by thermopile measurements using an
amplified thermocouple voltage difference signal which is measured
with the data acquisition system. The voltage difference is then
converted into a temperature difference. The uncertainty analysis
is conducted at a typical value for Vj and AV. The uncertainties
used for the voltage measurement are manufacturer data for the data
acquisition system. The voltage difference measurement with the
thermopile used type T thermocouple wire. A check of this
thermocouple wire against a NIST calibrated quartz thermometer
62
-------
resulted in absolute temperatures that were accurate within ±0.12 K
in the range of interest, which is well within the expected range
of ±1 K, However, since a thermopile measurement is used to
determine the temperature differences across the heat exchangers,
this test was conducted only to ensure the quality of the used
thermocouple wire. The importance of this random check lies in the
difference of the measured temperatures. Although the thermocouple
temperatures were up to 0.12 K different from the NIST calibrated
reference, the difference between the thermocouple temperatures did
not exceed 0.025 K, thus ensuring a high confidence in the here
calculated uncertainty of the temperature difference measurements.
The thermopile voltages are the sum of ten thermocouple pairs
across the measured device, i.e., the result is the ten times
amplified voltage difference between the thermocouple junctions on
each side of the heat exchanger. Due to this method the expected
error is much smaller than for the difference of two absolute
temperatures. The uncertainty can therefore be calculated as
follows;
&T=£(V1+&V) -f (Vt) . (VI)
V1 - 2mV ± (0.03 % + 6 \iV) (VII)
VL = 2 mV ± 0.0066 mV (VIII)
AF = 0.4 mV ± (0.03 % + 6 \lV) (IX)
AF = 0.4 mV ± (0.00072 mV) ¦ (X)
\
3.f (Fj+A F) 2 (FX+AF)
1 -dVt) + (¦ 1
3F,
0AF
2 SfiV )
dAF) + ( dVx )
3f(Fx+ AF)
9F,
dV% = 22.9738 — * 0.0066 mV - 0.15163 °C.
1 mV
8f(V +A V) ° n
-JL -dAF = 22.9738 — * 0.00072 °C = 0.016541 °C- •
8AF mV
df(Vx)
3F
dV, = 23.3824 — * 0.0066 mF = 0.15432 °C.
1 mV
EAt = 0.21698 °C.
e47= 2.3408 %
AT = 9 .26953 °C ± 0.21698 °C = 9 .26953 °C ± 2 .3408 % •
. (XI)
> (XII)
(XIII)
. (XIV)
. (XV)
(XVI)
(XVII)
63
-------
C.5 Heat, Transfer Fluid Specific Heat
The specific heat capacity of the heat "transfer fluid is
determined using the measured fluid temperature. Therefore, its
accuracy is dependent on the temperature measurement accuracy and
on the uncertainty of the regression curve. The function used to
determine the specific heat was interpolated from manufacturer
data. Figure CI shows the correlation function with respect to its
r~\
r\
V
-K
o>
w
-x
1
M
w
(d
(D
sz
u
If—
u
-------
Ec = * (M£dT)z + 0.008752 (XVIII)
\ oT
E,
P
[0.0034183 —* 1 K ) 2 + 0.008752 = 0.0093940 _J£tT_CXIX)
kg*Kz kg*K
ec = 0.2684 % (XX)
Cp = 3.618 , JCt7„ ± 0.0093940 kJ
kg*K kg*K
(XXI)
3.618 ± 0.2684 %
kg*K
C.6 Heat transfer fluid mass flow rate
The error analysis for the mass flow rate is based on the
manufacturer data for the calibrated device;
= 0.4%...... (XXII)
jfl = 0.08 ± 0 . 00032 -^2- =0.08 ± 0.4 % (XXIII)
sec sec sec
In order in ensure the accuracy of the measurements, the mass
flow metering device was calibrated prior to its use in the test
rig. Figures C2 and C3 show the calibration curves for both mass
flow meters. The values used for the calculation of the calibra-
tion curves lie within ±1.2% of the calibration curves, thus
showing a much higher uncertainty than expected. The value used
for the mass flow rate uncertainty in the following sections is
therefore chosen to be ±1.2%.
65
-------
D. 2
y=0.0015963+0,0S3BS6*x
cn
¦M
ed
0
Cf!
C/1
rd
£
0,05
•4
2
3
~
1
voltage C^O
caleu ieted I ntercept
Fig. C3: Mass flow meter calibration curve for condenser
heat transfer fluid measurements
66
-------
C.7 Heat Transfer Fluid Capacity Measurement
Q = A * Cp * AT = 2595.5 W. . .
(XXIV)
The capacity uncertainty is dependent on the uncertainties of the
contributing values used in the calculation. Therefore:
E,
6 \
(ih * c * 2 + (c_ * At * -i^dm)2
p a AT p dih
oc
+ (m * AT * -x—£dc„)
dcp p
(XXV)
40 "
(0.08
kg
sec
* 3500 , J * 0.013844 K)*
kg*K
+ (3500
kg*K
* 9.2695 /C ~ 0.00096 -^L)2
sec
+ (0.08 * 9 .2695 K * 9 .3940 , ^ )2
sec icg*JT
= \/ (3 . 8764 W}2 + (31.146 P^) 2 + (6.9662 W) 2
= 32.150 W
Sa =
32.150 W
* 100 = 1.2387 %
(XXVI)
(XXVII)
(XXVIII)
2595.5 W
d = 2595.5 W± 32.150 W = 2595.5 W± 1.2387 %• ...... (XXIX)
C.8 System Capacity Measurement
Since an indirect method is used to measure the system capacity,
the absolute result for the capacity is the heat transfer fluid
capacity. However, the error for the system capacity is greater
than that of the heat transfer fluid capacity since there is heat
loss in the condenser and heat gain in the evaporator. Here the
analysis is performed for a sample evaporator capacity under test
condition 1A and for the condenser capacity under test condition
ID. This is done because these are the capacities that are used to
calculate the system COP.
= 0 ± ( (Q2oss) eva + * 0) (XXX)
67
-------
Qcon-0± i ^ ~ (-jSL) . Q) (XXXI)
&l«,s - "J * a * (Tjx - T„) (XXXII)
with: X = 0.036 ——
m*K
5=0.0128 m
A-1.5JH* (XXXIII)
(Tmva - TJ = 2.5 K
(Tcon - TJ =5.5 K
Mioss) eva = Q° * 1-5 *2.5 W= 10.55 W. ....... (XXXIV)
(Qloss) - „° * J3^6' * 1.5 * 5.5 W = 23.20 W, . . . . . . . . (XXXV)
loss eon 0.0128
Qeva = 3520 W ± ( 10.55 W + 0.012387 * 3520 W)
= 3520 V ± 54.152 V * * (XXXVI)
= 3520 W± 1.5384 %
Qcon - 197 0 W± (23.20 W ± 0.012387 * 197 0 W)
= 1970 W ± 47.602 W (XXXVII)
= 1970 W * 2.4164 %
C.9 Compressor Power Measurement
The compressor power is measured with a torque and speed meter
which provides the data acquisition system with analog signals.
The device was manufacturer calibrated and is accurate within the
used uncertainties. Since the torque and speed meter
68
-------
was new at the beginning of the test series the authors did not see
a need for recalibration. Therefore;
W = — * Tor * 2 * % = 900 W ±
60 w
withi Tor = 5 .73 Nm ± 1%
n = 1500 -4— ±5 1
• • • •
mm
m in
(XXXVIII)
E*
<|>W + i^-dTor)'
dn dTor
(XXXIX)
s,
- \
. 2 . , , dn)2 ~ (JL .2 .« »drox>2
\
5.73Nm^2*n*5 1 x 2
60
sec
min
mm
(XL)
1500
+ (
mm
60
sec
min
* 2 * % * 0. 0573 Nm)
Ea = 9.486 W.
= 1.0541 %
W = 900 V ± 9.486 V = 900 W i 1.0541 %
(XLI)
(XLII)
(XLIII)
C.10 COP Calculations
To calculate the uncertainty of the COP, the uncertainties of the
system capacity measurement and of the compressor power measure-
ments are used. The data used represent typical values for the
high temperature cooling condition (evaporator) and low temperature
heating condition (condenser).
COP
W
(XLIV)
E,
COP
d(4-)
w
3<
dw
¦dti) + (¦
Q_
w
¦dQ)
\
tr
. 2 *1.2
am + (4 do)
w
(XLV)
69
-------
(- 3520 W
% (900 W) 2
= 0.072938
* 9.4867 W)2 + ( -— * 54,152 W)
900 W
2
(XLVI)
( 1970 W
\ (900 m2
* 9.4867 W)2 + ( i— * 47.602 W)
900 W
2
(XLVII)
0.057705
(XLVIII)
(XLIX)
COPeva = 3.911 ± 0.072938 = 3.911 ± 1.8649 %
(L)
COP,
con
- 2.188 ± 0.057705 » 2.188 ± 2.6363 %
(LI)
C.11 Refrigerant Composition
In order to find the composition of the mixtures actually used in
the test rig, samples of the discharge gas are taken during each
test. That sample is then analyzed in a gas chromatograph. The
chromatograph reports areas that are proportional to the number of
moles of each substance contained in the injected sample. In order
to calculate the mole fraction of each component it is necessary to
establish the areas for the pure components. From tests and
measurements at NIST it is known that the calibration' areas are
reported within ±0.1% with this particular machine. The areas of
each component in a mixture sample are typically repeatable within
±0.3%. The error of the mole fraction is neglected in this
consideration. Therefore the uncertainty analysis for the mass
fraction X can be performed using typical area values of a
R32/R134a composition measurement:
70
-------
Ml til :
Aa = 750590
±
0.3% = 750590
±
2251.8
Ab = 1267100
±
0.3% = 1267100
±
3801.3
A«i.. = 1726967
±
0.1% « 1726967
±
1727 .0
^cal,4 = 2342467
±
0.1% = 2342467
±
2342.5
MWa = 52.024 —2—
mole
MWb = 102.030 —2—
^ mole
•^a ~ Ab'^cal.a'Asal,!^
Am
*cal,a
~ Mffm
A- A i,
2— * MW1 + £
A?al, a ^
Cdi , 1>
* Mv^
JCa = caJ»a _ 0 . 29 06 ± Ey
a Aa . Aa MMbx x*
2_ + ( £_ * S.)
^cal, a ^caj, b MWg
Ev
dX 2
ar^' +
ax, , , ax.
^Ctej2 + { a
aA,
0A
cal, a
¦dAcaJ.J2
+ (
~3T
3a,
'a
¦ da.
cal, Is
cal,b>
1 A. Al MWhv 1 A,
± *[ £_ + ( £_ * £) ] - i * S_
_ ^cai, a ^cai,a ^cai, jb •^eai. a -^cai, a
5Aa " f Aa ^ A, MN2
)]
^cal, a ^cal, b
2 ,7467 *10"7
71
-------
. i MWb.
o - ( ±— * £) ~. a
A^l,a = -l . 627 0 *10~7 (LVII)
Ab [ + ^ Ab ^ j j 2
^ca.1, a ^c&l, b
A. _ -^a j ^Jh» . _ ,Aa _
a -. * [ *— + ( * £ ) ] - *_ * [ - —a
_ ^cal.a Acal,a Acal.b ^^a A=al,a &cal,a
__ . — ^ ^ ~ ^ u¥ ^ ^ 1 (LVIII)
OAcal,a. r a j. t Ab «"h.,o
[__*— + (__£— * _^) ]:
-1.1938 *1Q~7
^cal.a ^cal.h MW&
Aa * r Ab + MWf
^ « Aeal'a Ac*I,l,2 ^ = 8 , 8011*10"* (LIX)
^^cal, b r ^a + / ^b ^ \ "I 2
¦^ea J, a \*l,b a
from equation LIII:
dA. = 2251,8
a
£% = 3801.3 (LX)
<»«!,. = "27.0
dAcaifb = 2342.5
Ex = ij (6 ,1850*10~4) 2 + (-6 .1847 *10-4) 2
+ (-2 . 0617 #10~4) 2 + (2 . 0617 *10*4} 2 (LXI)
= 9 .2199 *10~4
ex = 0.31727% (LXII)
72
-------
xa = 0.2906 ± 9 .2199 *10~4 = 0.2906 ± 0.31727%. ...... (LXIII)
At the end of the tests a quality assuring recalibration of the gas
chromatograph was conducted. As expected, the values for the areas
representing the mole numbers of a substance were well within the
0.1% of the calibration that was conducted at the beginning of the
mixture tests.
73
------- |