NIST Technical Note 1895
­

Heat Pump Test Apparatus for the 

Evaluation of Low Global Warming 

Potential Refrigerants 

Harrison Skye 

http://dx.doi.org/10.6028/NIST.TN.1895 
 

 

 

NIST Technical Note 1895 

Heat Pump Test Apparatus for the 

Evaluation of Low Global Warming 

Potential Refrigerants 

Harrison Skye 

Engineering Laboratory 

Energy and Environment Division 

This publication is available free of charge from: 

http://dx.doi.org/10.6028/NIST.TN.1895 

November 2015 

U.S. Department of Commerce 

Penny Pritzker, Secretary 

National Institute of Standards and Technology 

Willie E. May, Under Secretary of Commerce for Standards and Technology and Director 
 

 

 

 

 

Certain commercial entities, equipment, or materials may be identified in this 

document in order to describe an experimental procedure or concept adequately. 

Such identification is not intended to imply recommendation or endorsement by the 

National Institute of Standards and Technology, nor is it intended to imply that the 

entities, materials, or equipment are necessarily the best available for the purpose. 

National Institute of Standards and Technology Technical Note 1895 

Natl. Inst. Stand. Technol. Tech. Note 1895, 48 pages (November 2015) 

http://dx.doi.org/10.6028/NIST.TN.1895 

CODEN: NTNOEF 
Abstract
­

International efforts to reduce human-induced global warming include restrictions on the 

use of chemicals with a high global warming potential (GWP). The heating, ventilation, airconditioning, and refrigeration (HVAC&R) industry subsequently faces a phasedown of many 

commonly used hydrofluorocarbon (HFC) refrigerants that have a relatively high GWP. A new 

family of refrigerants known as hydrofluoroolefins (HFOs), including their mixtures with HFCs, 

show promise as replacements. The overall GWP impact of refrigerant use includes both the 

direct GWP related to inadvertent release of the fluid into the atmosphere, as well as an indirect 

GWP caused by emissions from the power source used to energize the associated HVAC&R 

equipment. In nearly all HVAC&R applications, the indirect emissions far outweigh the direct 

emissions, so the efficiency of candidate replacements must be carefully quantified to guide the 

selection of fluids that actually achieve a reduction in the overall GWP. 

A 3.4 kW (1 ton) heat pump test apparatus has been constructed and instrumented for 

measuring the cycle performance of low-GWP refrigerants; the description of that apparatus is 

the focus of this report. Details are described for the system components, instrumentation, data 

reduction, and uncertainty analysis. Results from baseline experiments with R134a were used to 

test and verify the apparatus and data reduction procedure. The data will be used to provide a 

relative comparison between the low-GWP refrigerants as quantified by metrics including 

coefficient of performance and volumetric capacity. 

Future tests with low-GWP refrigerants will be compared with those from baseline 

refrigerants R134a and R410A. Additionally, the test data will be used to verify a new NIST 

heat pump modeling tool, CYCLE_D-HX, which captures both thermodynamic and heat transfer 

processes in HVAC&R equipment. Refrigerants, refrigerant mixtures, and cycle configurations 

that are difficult and/or time consuming to test can be rapidly explored using the verified model. 

ii
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Acknowledgements
­

This apparatus was constructed by John Wamsley based on the initial design by David 

Yashar. Dr. Yashar also helped create the outline of the report, which was valuable for keeping 

the presentation of information organized and complete. I would like to acknowledge Vance 

Payne for his help in determining the operating procedures and target operating parameters. 

Patrick Goenner refined the instrumentation and collected early sets of data while visiting as a 

guest researcher from Germany. I would also like to thank Piotr Domanski and Riccardo 

Brignoli for reviewing data sets and working on the CYCLE_D-HX model. Thanks to Mark 

Kedzierski for thoughtful insight into the calibration and uncertainty analysis for the test rig. 

Tara Fortin, of the NIST Applied Chemicals and Materials Division, performed differential 

scanning calorimeter measurements of the heat transfer fluid specific heat; these data were 

critical for achieving an acceptable energy balance. Finally, I would like to thank the reviewers 

for their helpful suggestions, including Piotr Domanski and Amanda Pertzborn at NIST, and 

Steve Brown at The Catholic University of America. 

iii
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Table of Contents
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Abstract........................................................................................................................................... ii
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Acknowledgements........................................................................................................................ iii
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Table of Contents........................................................................................................................... iv
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List of Figures................................................................................................................................ vi
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List of Tables ................................................................................................................................ vii
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Nomenclature............................................................................................................................... viii
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1 Introduction ............................................................................................................................... 1
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2 Test Apparatus .......................................................................................................................... 3
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2.1 Components....................................................................................................................... 3
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2.2 Instrumentation ................................................................................................................ 11
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2.3 Operation ......................................................................................................................... 15
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2.3.1 Test Rig Control Parameters.................................................................................... 15
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2.3.2 Test Rig Operating Parameters ................................................................................ 15
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2.3.2.1	­ Operating Parameters 1 & 2: Evaporator and Condenser Saturation 

Temperature...................................................................................................... 15
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2.3.2.2	­ Operating Parameters 3, 4, & 5: Capacity and Heat Flux ................................ 17
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2.3.2.3	­ Operating Parameters 6 & 7: HTF-side Thermal Resistance Ratio ................. 18
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2.3.2.4	­ Operating Parameters 8 & 9: Subcool and Superheat ...................................... 18
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2.3.2.5 Operating Parameter 10: LLSL-HX included or bypassed .............................. 19
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3 Data Analysis.......................................................................................................................... 20
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3.1 Thermodynamic States .................................................................................................... 20
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3.2 Thermodynamic Performance ......................................................................................... 20
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3.2.1 Capacity ................................................................................................................... 20
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3.2.2 Compressor Power................................................................................................... 21
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3.2.3 Coefficient of Performance...................................................................................... 22
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3.2.4 Volumetric Capacity ................................................................................................ 22
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3.2.5 Compressor Efficiency............................................................................................. 22
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3.2.6 Superheat and Subcooling........................................................................................ 23
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3.2.7 LLSL-HX Effectiveness .......................................................................................... 23
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4 Validation and Baseline Tests with R134a ............................................................................. 25
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5 Conclusions and Future Work ................................................................................................ 28
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6 References............................................................................................................................... 29
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iv 
 

 

Appendix A : Uncertainty Analysis.............................................................................................. 31
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A.1 Symbols Used in Uncertainty Analysis........................................................................... 31
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A.2 General Remarks ............................................................................................................. 32
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A.3 Thermocouples with CJC Compensation ........................................................................ 32
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A.4 Thermocouples with Ice-Water Bath Compensation ...................................................... 33
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A.5 Thermopiles in Heat Transfer Fluid ................................................................................ 34
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A.6 Evaporator Heat Transfer Fluid Specific Heat ................................................................ 36
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A.7 Condenser Heat Transfer Fluid Specific Heat................................................................. 37
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A.8 Evaporator and Condenser Insulation Conductance ....................................................... 37
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A.9 Uncertainty Analysis Software........................................................................................ 39
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A.10 Moving Window and Steady State Uncertainty.............................................................. 40
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A.11 Uncertainty Results Summary......................................................................................... 41
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A.11.1 Uncertainty Results Summary: Cooling .................................................................. 41
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A.11.2 Uncertainty Results Summary: Heating................................................................... 43
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Appendix B : Instrumentation Schematic with DAQ Number Convention ................................. 46
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Appendix C : Microfin Tube Surface Area................................................................................... 48
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v 
 

 

 

 

 

 

 

 

 

List of Figures
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Figure 2-1: Schematic of MBBHP test apparatus...................................................................... 5 

Figure 2-2: Pictures of MBBHP apparatus................................................................................ 6 

Figure 2-3: Pictures of MBBHP compressor (a) without and (b) with safety cage................... 7 

Figure 2-4: Schematic showing the tube numbering convention in the (a) condenser and (b)
­

evaporator. .............................................................................................................. 8 

Figure 2-5: Schematics of annular heat exchanger including (a) refrigerant tube lengths, (b)
­

cross section of annular heat exchanger, (c) detail cross-section of microfin tube, 

and (d) helix angle of microfins.............................................................................. 9 

Figure 2-6: Capacity and heat flux for baseline R134a tests for (a) evaporator and (b)
­

condenser .............................................................................................................. 18 

Figure 3-1: Temperature profile in (a) evaporator, showing superheat and (b) condenser, 

showing subcooling .............................................................................................. 23 

Figure 4-1: Energy imbalance in the evaporator and condenser ............................................. 25 

Figure 4-2: Heat pump cycle pressures on the (a) high pressure side and (b) low pressure
­

side ........................................................................................................................ 26 

Figure 4-3: Performance metrics including (a) compressor power and (b) mass flow and mass
­

flux ........................................................................................................................ 26 

Figure 4-4: Performance metrics including (a) COP, (b) capacity, (c) volumetric capacity, and 

(d) compressor isentropic and volumetric efficiency ............................................ 27 

Figure 4-5: Flow regime map for microfin tube (a) evaporation and (b) condensation .......... 28 

Figure A-1: Temperature calibration for CJC compensated thermocouples where (a) data are
­

Figure A-4: Contours of thermopile (a) uncertainty and (b) relative contribution to 

Figure A-8: Uncertainty in temperature due to fluctuations at steady state for (a) evaporator
­

divided by each of the 5 calibrations and (b) all data are combined..................... 33 

Figure A-2: Ice-water bath referenced thermocouple calibration data..................................... 34 

Figure A-3: Thermopile calibration data .................................................................................. 35 

uncertainty............................................................................................................. 36 

Figure A-5: Specific heat measurements for evaporator HTF.................................................. 37 

Figure A-6: Specific heat of condenser HTF (water) ............................................................... 37 

Figure A-7: Insulation conductance for evaporator and condenser.......................................... 39 

outlet and (b) evaporator inlet............................................................................... 41 

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List of Tables
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Table 2-1: Components for MBBHP rig ................................................................................... 10 

Table 2-2: Refrigerant microfin tube specifications.................................................................. 11 

Table 2-3: List of MBBHP instruments (SI units) .................................................................... 13 

Table 2-4: List of MBBHP instruments (I-P units)................................................................... 14 

Table 2-5: List of the Control Parameters and Operating Parameters ...................................... 15 

Table 2-6: Evaporator and condenser fluid temperatures for standard rating tests................... 17 

Table 3-1: Measurements and equations used to define the thermodynamic states.................. 20 

Table A-1: Condenser & evaporator thermopile polynomial coefficients................................. 35 

Table A-2: Evaporator HTF specific heat polynomial coefficients ........................................... 37 

Table A-3: Uncertainty of refrigerant-side evaporator cooling capacity ................................... 42 

Table A-4: Uncertainty of HTF-side evaporator cooling capacity ............................................ 42 

Table A-5: Uncertainty of cooling COP .................................................................................... 43 

Table A-6: Uncertainty of refrigerant-side condenser heating capacity .................................... 44 

Table A-7: Uncertainty of HTF-side condenser heating capacity.............................................. 44 

Table A-8: Uncertainty of heating COP..................................................................................... 45 

vii
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Nomenclature
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Symbol Units Definition 

A m2 Area 

Atotal m2 Total combined heat transfer area from evaporator and condenser 

COPheat -- Coefficient of Performance - Heating 

COPcool -- Coefficient of Performance - Cooling 

Cp kJ kg -1 K-1 Specific heat 

do mm Microfin tube outer diameter 

di mm Microfin tube inner diameter 

Dcomp m3 Compressor displacement 

h kJ kg -1 Specific enthalpy 

L m Tube length 

mɺ kgs -1 Mass flow rate 

N Hz Compressor frequency 

Nf -- Number of microfins on inner circumference of refrigerant tube 

NT -- Number of heat exchanger tubes 

P kPa Pressure 

∆P kPa Pressure difference 

R K kW -1 Thermal resistance 

Rtotal K kW -1 Total resistance in heat exchanger including both fluids and tube wall 

Qɺ kW Energy transfer 

Qɺ kW Energy transfer in LLSL-HX, on the vapor side LLSL v,

Qɺ , kW Maximum possible energy transfer in LLSL-HX LLSL max

T °C Temperature 

∆T K Temperature difference 

tb mm Microfin tube bottom wall thickness 

sf mm Spacing between microfins 

s kJ kg -1 K-1 Specific entropy 

SC K Condenser subcooling 

SH K Evaporator superheat 

UA W K -1 Thermal conductance 

v m3 kg-1 Specific volume 

VCC kJ m -3 Volumetric Cooling Capacity 

VHC kJ m -3 Volumetric Heating Capacity 

Wɺ

comp kW Compressor power, computed using enthalpy change of refrigerant 

Wɺ kW Compressor power, computed using torque and speed of driveshaft shaft 

x -- Vapor quality (mass of vapor divided by total mass of fluid) 

viii
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c 

Greek 

Symbol Units Definition 

α ° Microfin helix angle about tube axis 

β ° Microfin angle 

ε LLSL -- Effectiveness of the liquid-line suction-line heat exchanger 

ηs -- Compressor isentropic efficiency 

ηv -- Compressor volumetric efficiency 

ρair kgm -3 Density of air 

τ N m Compressor shaft torque 

Subscript Definition 

active Active refrigerant heat exchanger tubes
­

air Air
­

amb Ambient air surrounding test apparatus
­

avg Average
­

Condenser 

e Evaporator 

HTF Heat Transfer Fluid 

in Inlet 

ins Insulation 

inactive Inactive refrigerant heat exchanger tubes 

out Outlet 

ref Refrigerant 

1 to 11 Refrigerant thermodynamic states as defined in Figure 2-1 

ix 
 

 

Abbreviation Definition 

AHRI Air-Conditioning, Heating, and Refrigeration Institute 

ANSI American National Standards Institute 

CFC Chlorofluorocarbon 

COP Coefficient of Performance (W of capacity per W of electric input) 

CYCLE_D-HX A NIST heat pump cycle simulation model currently under development. 

The simulation captures both thermodynamic and transport processes in the 

cycle. 

DAQ Data Acquisition 

DOE Department of Energy (United States) 

EER Energy Efficiency Ratio (Btu of capacity per W of electric input) 

EOS Equation of State 

GWP Global Warming Potential 

HFC Hydrofluorocarbon 

HFO Hydrofluoroolefin 

HTF Heat Transfer Fluid 

HVAC&R Heating, Ventilation, Air-Conditioning, and Refrigeration 

LLSL-HX Liquid-Line Suction-Line Heat Exchanger 

MBBHP Mini Breadboard Heat Pump 

NIST National Institute of Standards and Technology (United States) 

PID Proportional Integral Derivative controller 

RPM Revolutions per Minute (compressor shaft) 

x 
 

 

 

1 Introduction
­

Concerns about the environmental impacts of global warming (i.e., global climate change) 

are driving an effort to limit anthropogenic sources of atmospheric pollutants that trap longwave 

radiation emitted from the earth’s surface. The chlorinated- and fluorinated-hydrocarbon 

working fluids (i.e., refrigerants) employed by the heating, ventilation, air-conditioning, and 

refrigeration (HVAC&R) industry exhibit a particularly large global warming potential (GWP) 

with 100-year GWP values hundreds to thousands of times larger than an equivalent mass of 

carbon dioxide (Solomon, et al., 2007, p. 212). In the European Union, the F-gas regulation (EU, 

2014) mandates a phase-down of hydrofluorocarbons (HFCs) with high GWP to 21 % of the 

average levels from years 2009 through 2012 by 2030. The current North American proposal 

would amend the Montreal Protocol to limit HFC consumption to have a weighted GWP of 15 % 

of average levels from years 2011 through 2013 by 2036 (EPA, 2015). 

Major efforts are underway to identify alternative refrigerants with a lower GWP. Chemical 

manufacturers are proposing halogenated olefins (e.g., hydrofluoroolefins, or HFOs), which offer 

substantially lower GWP due to short atmospheric lifetimes. For some HFOs, the lower GWP 

comes at the cost of an increase in flammability. The Air-Conditioning, Heating, and 

Refrigeration Institute (AHRI) is leading a collaborative effort to evaluate the drop-in and softoptimized performance of low-GWP alternatives largely consisting of HFOs and HFO/HFC 

blends (AHRI, 2015). The National Institute of Standards and Technology (NIST) investigated 

the maximum thermodynamic potential expected for working fluids by optimizing the 

parameters governing the Equations of State (EOS) (Domanski et al., 2014). In a companion 

NIST study, a coarse filter criteria was developed (based on the atomic elements, GWP, toxicity, 

flammability, critical temperature, and stability) to sift through the 100 million chemicals 

currently listed in the public-domain PubChem database for possible low-GWP refrigerant 

candidates, 1234 of which were identified (Kazakov et al., 2012). The candidate list was further 

refined to 62 by limiting the critical temperature to the range 300 K to 400 K, which is typical for 

fluids in HVAC&R equipment (McLinden et al., 2014). 

The remaining 62 candidates, and their mixtures, must be evaluated using more stringent 

criteria that consider detailed performance in HVAC&R equipment. In particular, the criteria 

must include the cycle efficiency of the fluid since the carbon dioxide emissions from the power 

source (i.e., “indirect emissions”) far outweighs the direct GWP impact (i.e., the GWP from the 

release of the fluid into the atmosphere) of the working fluid in nearly all HVAC&R equipment. 

A cycle modeling tool, CYCLE_D-HX, is being developed at NIST to evaluate the efficiency of 

low-GWP refrigerants and refrigerant mixtures. The model expands upon the thermodynamic 

considerations from CYCLE_D (Brown et al., 2011) by including transport phenomena. In this 

way, appropriate penalties can be applied to refrigerants whose thermodynamic properties are 

favorable but exhibit poor performance in HVAC&R systems because of large pressure drops or 

poor heat transfer. The consideration of transport processes represents a lesson learned from the 

previous effort to find replacements for chlorofluorocarbon (CFC) fluids that were found to 

cause stratospheric ozone destruction. R410A, the dominant replacement for the CFC R22, was 

initially considered to be a significantly inferior replacement based on thermodynamic properties 

alone. However, later practice showed that the lower pressure drop (due to high operating 

pressure) and superior heat transfer of R410A could be exploited with optimized heat exchangers 

to yield performance competitive with R22. Including transport processes also allows for the 

1 
 

 

 

 

 

 

optimization of mass flux; the number of parallel heat exchanger tube circuits are adjusted to 

balance performance enhancement/degradation of increased heat transfer/pressure drop with 

higher mass flux. The more sophisticated capabilities of the new model will allow for more 

accurate identification of optimal low-GWP fluids using computational methods. 

A heat pump test apparatus has been constructed for the purpose of generating cycle 

performance data to compare low-GWP refrigerants and to tune and verify the new NIST model; 

that test apparatus is the focus of this report. The apparatus is highly configurable, which allows 

the system to operate with a wide variety of working fluids as if tailored to each one. 

Specifically, the apparatus features a variable-speed compressor, variable-area heat exchangers, 

and a manually adjusted throttle valve. These adjustable components allow for operation at ɺ constant capacity per total combined evaporator and condenser conductance (Q UA ) across total

all test fluids, a condition required for equitable comparison of refrigerants (as well as fixed heat 

transfer fluid inlet/outlet temperatures) (McLinden et al., 1987). In practice, it is difficult to hold 

Q UA ɺ constant because the conductance depends on the refrigerant side heat transfer total

coefficient, which changes for each fluid and operating condition. Rather, a fixed capacity per 

total heat exchanger area ( Qɺ A ) is used as a close approximation. For example, with a fixed total 

heat exchanger area, Qɺ A is held constant from R134a (low volumetric capacity) to R410A total 

(high volumetric capacity) by decreasing the compressor speed. If adjustment of compressor 

speed is insufficient to maintain constant Qɺ A between refrigerants, the heat exchanger area total 

can be adjusted as well. The adjustability and relatively small capacity of 1.7 kW to 3.5 kW (0.5 

ton to 1 ton) provide the etymology for the apparatus name, the Mini Breadboard Heat Pump 

(MBBHP). Measurements of cycle performance with legacy high-GWP HFCs and novel lowGWP fluids in the MBBHP will provide a rich data set to test the predictive capability of the new 

NIST model. 

This report describes the construction and operation of the test apparatus, the data analysis 

used to quantify performance, an uncertainty analysis of the performance metrics, and baseline 

and validation results using refrigerant R134a. 

2 
 

 

 

 

 

 

 

 

 

 

2 Test Apparatus
­

2.1 Components 

The test apparatus, shown in Figure 2-1 through Figure 2-3, was designed to emulate 

refrigerant-side operating conditions of typical HVAC&R equipment. (Figure B-1 and Figure B2 in Appendix B show the instrument numbering convention used in the raw data files). The rig 

design is partially based on one used in a previous NIST effort to investigate alternatives to 

CFCs (Pannock et al., 1991). The refrigerant circuit includes a variable-speed compressor 

powered by an electric motor and inverter, where the speed controls cooling/heating capacity. 

The evaporator and condenser, shown in detail in Figure 2-4 and Figure 2-5 (a) and (b), are 

single-circuit (i.e., there are no parallel tube branches) annular heat exchangers where the 

refrigerant is on the inner tube and the liquid heat transfer fluid (HTF) is in the annular space. 

As shown in Figure 2-5(c), the refrigerant tube is smooth on the outside and enhanced with rifled 

microfins on the inside. The size of the heat exchangers can be adjusted by changing the number 

of active refrigerant tubes; the valves on the side of the heat exchanger allow for up to 10 (in 

increments of 2) of the 20 refrigerant tubes to be bypassed (this section of the rig is referred to as 

the “bypass valve header”). When operating these bypass valves, care is taken to avoid trapping 

subcooled liquid in a closed section; a section completely filled with subcooled liquid will 

experience a catastrophic increase in pressure if the temperature rises enough to vaporize the 

refrigerant. A liquid-line suction-line heat exchanger (LLSL-HX) can be included or bypassed 

as needed. Each of three heat exchangers (evaporator, condenser and LLSL-HX) are arranged in 

counterflow configuration. Two sizes of manually controlled Vernier-handled needle valves are 

used to throttle the refrigerant and can be used individually or in tandem; the needle valve sizes 

were selected using a correlation for short tubes (Kim et al., 2005). A filter-dryer is used to 

control contaminants, and an oil separator is used to ensure proper oil return to the compressor 

and to minimize oil circulation in the heat exchangers. The system refrigerant charge is adjusted 

using two Shrader access valves, where refrigerant is removed from or added to the 

liquid/suction line valves, respectively. Two 40 W heaters preheat the compressor before startup 

to avoid foam in the oil (the heaters are turned off after the discharge temperature has warmed to 

about 60 °C). A safety cage was constructed around the compressor because during cooling tests 

with high pressure fluids such as R410A, the compressor will be operated slightly above the 

manufacturer’s specified maximum pressure of 2600 kPa (363 psig). Finally, a series of sight 

glasses provide visual confirmation of fully subcooled liquid exiting the condenser, vapor only 

(i.e., no liquid) entering the compressor suction port, proper oil level in the compressor, and 

proper oil return from the oil separator. Detailed component specifications are listed in Table 

2-1. The microfin tube parameters are shown in Figure 2-5 and Table 2-2, and Appendix C 

shows how the microfin surface area was calculated. 

The HTF circuits include a water heater/chiller to respectively control the evaporator and 

condenser HTF inlet temperatures, where the heater/chiller contains a pump for circulating the 

fluid. The evaporator HTF is a potassium formate and water solution with freeze protection 

to -40 °C (-40 °F), whereas water is used for the HTF in the condenser. Flow control provided 

by three valves in each HTF circuit allows for a wide range of flowrates: two in-line valves are 

used where the larger/smaller (gate/needle) valve provides gross/fine flow adjustment and the 

third valve bypasses the heater (or chiller) for flow and pressure control. Air bubbles in the 

condenser HTF circuit are controlled by drawing from the bottom of the vented chiller reservoir 

3 
tank. The evaporator HTF circuit is pressurized with an expansion tank, so air bubbles can be 

removed by venting trapped air from local high spots using Shrader valves. Experience 

operating the test apparatus has shown that the vented reservoir tank (in the chiller) is a superior 

solution for minimizing air bubbles. The temperature of the fluid delivered to the heat 

exchangers is regulated using proportional-integral-derivative (PID) controllers. 

4 
Figure 2-1: Schematic of MBBHP test apparatus 

5 
Figure 2-2: Pictures of MBBHP apparatus 

6
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(a) (b)
­

Figure 2-3: Pictures of MBBHP compressor (a) without and (b) with safety cage.
­

7
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(a)	­ (b) 

Figure 2-4:	­ Schematic showing the tube numbering convention in the (a) condenser 

and (b) evaporator. 

8
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(a)
­

(b)	­ (c) (d) 

Figure 2-5:	­ Schematics of annular heat exchanger including (a) refrigerant tube 

lengths, (b) cross section of annular heat exchanger, (c) detail crosssection of microfin tube, and (d) helix angle of microfins. 

9
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Table 2-1: Components for MBBHP rig 

Component Characteristics 

Chiller Cooling capacity: 2500 W (0.7 ton) 

Compressor Type: 

Cylinders: 

Speed: 

Displacement rate: 

Displacement: 

Max pressure: 

Reciprocating 

2 

8.3 Hz to 66.7 Hz 

(500 RPM to 4000 RPM) 

1.31 m3

h-1 to 10.81 m3h-1 

7.165 cm3 

2604 kPa (363 psia) 

Electric motor Maximum power: 3.73 kW (5 hp) 

Expansion tank (HTF) Size: 19 liter (5 gallon) 

Filter (HTF) Filter size: 50 μm 

Filter-dryer (refrig.) Line size: 9.5 mm (0.375 in) 

Inverter Voltage: 

Maximum power: 

230 VAC, 3-phase 

3.73 kW (5 hp) 

Heater Heating capacity: 10 kW (2.8 ton) 

Heat exchanger 

(evaporator/condenser) 

Number of tubes: 

Tube active length: 

Return bend length: 

Outer tube ID: 

Outer tube OD: 

Outer tube material: 

Inner tube: 

10 to 20, increments of 2 

560 mm (22 in) 

83 mm (3.25 in) 

25.4 mm (1 in) 

26.8 mm (1.125 in) 

Copper 

See Table 2-2 

Liquid-line suction-line heat 

exchanger (LLSL-HX) 

Nominal capacity: 

Shell O.D.: 

Type: 

370 W (0.5 hp) 

51 mm (2 in) 

Corrugated (liquid side) 

Needle valves (refrig.) Valve 1 

Orifice: 

Flow coeff. (Cv): 

Valve 1 

1.42 mm (0.056 in)* 

0.03* 

Valve 2 

Orifice: 

Flow coeff. (Cv): 

Valve 2 

3.25 mm (0.128 in)* 

0.16* 

Needle valves (HTF) Orifice: 

Flow coeff. (Cv): 

6.4 mm (0.25 in)* 

0.73* 

Oil separator Diameter: 

Height: 

Capacity: 

102 mm (4 in) 

356 mm (14 in) 

10 to 17 kW (3 to 5 ton) 

Pressure relief valve Cracking pressure: 800 psig 

*Note: valve orifice sizes and flow coefficients (Cv) can vary between 0 and the listed 

maximum values 

10 
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

Table 2-2: Refrigerant microfin tube specifications 

Parameter Symbol Value Uncertainty 

Inside diameter di 8.46 mm (0.333 in) Unknown 

Outside diameter do 9.52 mm (0.375 in) ±0.050 mm (±0.002 in) 

Bottom wall thickness tb 0.33 mm (0.013 in) ±0.025 mm (±0.001 in) 

Fin height e 0.20 mm (0.008 in) ±0.025 mm (±0.001 in) 

Spacing between fins sf 0.22 mm (0.009 in) Unknown 

Number of fins Nf 60 --

Fin angle β 60° Unknown 

Helix angle α 18° 2° 

Internal surface area Ai/L 0.0424 m (1.67 in) 

per unit tube length 

Material Copper 

2.2 Instrumentation 

Refrigerant-side instrumentation is selected to capture the key thermodynamic state points in 

the cycle. Specifications for the instruments (including instrument uncertainty) in both the 

refrigerant and HTF circuits are listed in Table 2-3 (Table 2-4 shows I-P units). The pressure 

transducers and in-stream thermocouple probes are numbered in Figure 2-1 according to the 

eleven states computed in Section 3.1; the thermodynamic state number convention is also 

consistent with that used in the new CYCLE_D-HX model. The thermocouple probes are 

referenced to a cold junction in an ice-water bath. Differential pressure transducers quantify the 

pressure drop in the condenser, evaporator, and suction line. The location of the heat exchanger 

pressure taps has been selected at the bottom of the bypass valve header, where the pressure is 

equal to that of the refrigerant leaving the heat exchanger regardless of how many tubes are 

bypassed. A coriolis flowmeter measures the refrigerant mass flow and density, and provides 

confirmation of fully subcooled liquid in the liquid line (a two-phase mixture will exhibit large 

oscillations in measured flowrate). The compressor shaft power is captured by the in-line torque 

meter and tachometer. 

On the HTF fluid side, energy transfers are measured in the evaporator and condenser to 

provide verification of the refrigerant-side calculations. The mass flow and temperature 

difference in the heat exchangers are measured with coriolis meters and thermopiles (in 

thermowells), which are combined with the fluid heat capacity to compute energy transfer. 

Thermocouples inserted into the thermowells with the thermopiles provide the measure of 

absolute temperature required for applying the thermopile calibration. Additionally, the 

thermocouples provide verification of the thermopile temperature difference measurement. 

Figure 2-1 and Figure 2-4 also show the tube-surface-mounted thermocouples used to 

measure both refrigerant and HTF temperature profiles in the evaporator and condenser. As 

shown in Section 3.2.6, these temperature profiles are useful for determining refrigerant tube 

11 
 

 

 

number where the phase transitions (superheat, subcooling, two-phase) occur. The refrigerantand HTF-side sensors are mounted on the return bends of each inner/outer tube, respectively. 

The instruments are scanned once per minute by the data acquisition (DAQ) system, and 

when steady state is achieved, the measurements are calculated as the average of a moving 

window of 30 readings (i.e., the average over the last 30 minutes). The measurements inevitably 

fluctuate during steady state, which contributes to the uncertainty of the measurement. This 

“steady state” uncertainty has been quantified for a nominal test condition and is listed in Table 

2-3. The total measurement uncertainty (last column of Table 2-3) is found by adding the 

instrument error and the steady state error in quadrature. A detailed discussion of the selection 

of the steady state moving window size and computation of steady state uncertainty is shown in 

Appendix A.10. 

12 
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

Table 2-3: List of MBBHP instruments (SI units) 

Instrument Range 

Instrument 

Uncertainty1 

Steady State 

Uncertainty1 

Total 

Uncertainty1,2 

Coriolis meter (HTF) Mass flow: 

(0 to 756) gs -1 

Mass flow 3: 

±0.1% rdg. 

Mass flow: 

±0.23 gs -1 

Mass flow: 

±0.26 gs -1 

Density: 

(0 to 3) g cm-3 

Density: 

±0.001 gcm-3 

Density: 

±0.0001 gcm-3 

Density: 

±0.001 gcm-3 

Coriolis meter (refrigerant) Mass flow: Mass flow 4: Mass flow: Mass flow: 

(0 to 60)5 gs -1 ±0.1% rdg. ±0.01 gs-1 ±0.016 gs -1 

Density: Density: Density: Density: 

(0 to 3) g cm-3 ±0.0005 gcm-3 ±0.0001 gcm-3 ±0.0005 gcm-3 

Data acquisition system (-10 to 10) V ±1 μV -- ±1 μV 

Press. transducer (diff.) – condenser (0 to 172) kPa ±0.7 kPa ±0.8 kPa ±1.3 kPa 

Press. transducer (diff.) – evaporator (0 to 345) kPa ±0.1 kPa ±0.8 kPa ±0.8 kPa 

Press. transducer (diff.) – suction (0 to 69) kPa ±0.01 kPa ±0.3 kPa ±0.3 kPa 

Press. transducer (high press.) (0 to 3447) kPa ±3.4 kPa ±0.7 kPa ±3.5 kPa 

Press. transducer (low press.) (0 to 1724) kPa ±3.4 kPa ±0.7 kPa ±3.5 kPa 

Tachometer (0.17 to 1667) Hz ±0.0167 Hz ±0.005 Hz ±0.0173 Hz 

(10 to 99 999) RPM ±1 RPM ±0.3 RPM ±1.04 RPM 

Thermocouple (in-stream probe, ice ref.) (-10 to 100) °C ±0.08 °C ±0.06 °C ±0.09 °C 

Thermocouple (surface-mount, CJC) (-10 to 60) °C ±0.6 °C ±0.04 °C ±0.6 °C 

Thermopile5 (1 to 10) K ±0.015 K ±0.013 K ±0.0153 K 

Thermopile voltage (-10 to 10) V ±1 μV ±6.5 μV ±7 μV 

Torque meter (0 to 40) N m ±0.33 Nm ±0.005 Nm 0.33 Nm 

1 95% confidence interval (k = 2)
­

2 % rdg values have been applied to nominal values
­

3 Maximum of 0.1% and (z.s./flowrate)*100 %, where z.s. is zero stability = 0.0491 g s -1 .
­

4 Maximum of 0.1% and (z.s./flowrate)*100 %, where z.s. is zero stability = 0.0075 g s -1 .
­

5 For temperature differences in 1 K to 10 K range. Outside range, uncertainty is higher. See Appendix A.5 for more
­

detail. 

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Table 2-4: List of MBBHP instruments (I-P units) 

Instrument Range 

Instrument 

Uncertainty1 

Steady State 

Uncertainty1 

Total 

Uncertainty1,2 

Coriolis meter (HTF) Mass flow: 

(0 to 100) lb min-1 

Mass flow 3: 

±0.1% rdg. 

Mass flow: 

±0.03 lb min-1 

Mass flow: 

±0.035 lb min-1 

Density: 

(0 to 187) lb ft-3 

Density: 

±0.062 lb ft-3 

Density: 

±0.006 lb ft-3 

Density: 

±0.062 lb ft-3 

Coriolis meter (refrigerant) Mass flow: Mass flow 4: Mass flow: Mass flow: 

(0 to 79.9) lb min-1 ±0.1% rdg. ±0.0015 lb min-1 ±0.002 lb min-1 

Density: Density: Density: Density: 

(0 to 187) lb ft-3 ±0.062 lb ft-3 ±0.006 lb ft-3 ±0.062 lb ft-3 

Data acquisition system (-10 to 10) V ±1 μV -- ±1 μV 

Press. transducer (diff.) – condenser (0 to 25) psi ±0.1 psi ±0.1 psi ±0.18 psi 

Press. transducer (diff.) – evaporator (0 to 50) psi ±0.015 psi ±0.1 psi ±0.1 psi 

Press. transducer (diff.) – suction (0 to 10) psi ±0.0015 psi ±0.04 psi ±0.04 psi 

Press. transducer (high press.) (0 to 500) psia ±0.5 psi ±0.1 psi ±0.51 psi 

Press. transducer (low press.) (0 to 250) psia ±0.5 psi ±0.1 psi ±0.51 psi 

Tachometer (0.17 to 1667) Hz ±0.0167 Hz ±0.005 Hz ±0.0173 Hz 

(10 to 99 999) RPM ±1 RPM ±0.3 RPM ±1.04 RPM 

Thermocouple (in-stream probe, ice ref.) (14 to 212) °F ±0.15 °F ±0.11 °F ±0.18 °F 

Thermocouple (surface-mount, CJC) (14 to 140) °F ±1.1 °F ±0.08 °F ±1.1 °F 

Thermopile5 (1.8 to 18) °F ±0.027 °F ±0.023 °F ±0.04 °F 

Torque meter (0 to 350) in lb ±2.9 in lb ±0.045 in lb ±2.9 in lb 

1 95% confidence interval (k = 2) 

2 % rdg values have been applied to nominal values 

3 Maximum of 0.1% and (z.s./flowrate)*100 %, where z.s. is zero stability = 0.0065 lb min-1 

4 Maximum of 0.1% and (z.s./flowrate)*100 %, where z.s. is zero stability = 0.001 lb min-1 

5 For temperature differences in 1.8 °F to 18 °F range. Outside range, uncertainty is higher. See Appendix A.5 for more 

detail. 

14 
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

2.3 Operation 

2.3.1 Test Rig Control Parameters 

There are ten Control Parameters (i.e., settings of adjustable components) listed in Table 2-5 

that the MBBHP operator sets to achieve a desired set of Operating Parameters. Each Control 

Parameter primarily governs one or two Operating Parameters; the correspondence is shown in 

the table. 

Table 2-5: List of the Control Parameters and Operating Parameters 

# Operating Parameter Operating Parameter Values Control Parameter 

1 Evaporator saturation temperature See Table 2-6 Heater setpoint (HTF inlet temp.) 

2 Condenser saturation temperature See Table 2-6 Chiller setpoint (HTF inlet temp.) 

3 Capacity/heat flux 1760 W to 3520 W Compressor speed 

(0.5 ton to 1 ton) 

4 Evaporator heat flux 2000 W m-2 to 12 000 W m-2 Number of evaporator tubes 

(600 Btu h-1 ft-2 to 3800 Btu h-1 ft-2) 

5 Condenser heat flux 2000 W m-2 to 12 000 W m-2 Number of condenser tubes 

(600 Btu h-1 ft-2 to 3800 Btu h-1 ft-2) 

6 Condenser HTF-side resistance 50 % to 70 % of overall resistance Condenser HTF flowrate 

7 Evaporator HTF-side resistance 70 % to 85 % of overall resistance Evaporator HTF flowrate 

8 Subcool/ Superheat ≤ 2 tubes1 / ≤ 1 tube2 Refrigerant needle valve setting 

9 Subcool/ Superheat ≤ 2 tubes1 / ≤ 1 tube2 Refrigerant charge 

10 LLSL-HX included/bypassed Included/bypassed LLSL-HX valves 

1

achieved with subcooling of 2 K to 3 K (1.8 °F to 3.6 °F) 

2

achieved with superheat of 3 K to 6 K (5.4 °F to 10.8 °F) 

2.3.2 Test Rig Operating Parameters 

This section describes each of the target Operating Parameters, and how they are set using 

the Control Parameters. Note that the headings for Sections 2.3.2.1 through 2.3.2.5 include the 

Operating Parameters numbers (1 to 10) as listed in Table 2-5. 

2.3.2.1 Operating Parameters 1 & 2: Evaporator and Condenser Saturation Temperature 

The U.S. Department of Energy (DOE) requires heat pumps and air conditioners to be rated 

at conditions prescribed in the ANSI/AHRI 210/240-2008 (AHRI, 2008) standard. The standard 

specifies temperatures of the HTF (air, in the standard) entering the evaporator and condenser. A 

somewhat different, but derivative, method is employed here; the HTF inlet temperatures are set 

to achieve refrigerant saturation temperatures expected under ANSI/AHRI Cooling A, Cooling B 

and Heating H1 Rating Tests for the baseline refrigerant R134a. 

The target saturation temperatures are computed first for the Cooling A test where the 

assumed Coefficient of Performance (COPcool) is 4.4 (energy efficiency ratio, EER, is 15). The 

15 
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

ANSI/AHRI standard is not prescriptive about airflow, so typical values per unit evaporator ɺ capacity are used: for the evaporator (Vɺ , Qevap ), 0.189 m3 s-1 kW-1 (400 ft3min-1 ton-1), and air evap ɺ for the condenser (Vɺ , Qevap ), 0.378 m3 s-1 kW-1 (800 ft3min-1 ton-1). Inlet air temperatures air cond 

(Tair,in) prescribed by the ANSI/AHRI standard are listed in Table 2-6; the temperature change 

across the evaporator and condenser (∆Tair,evap, ∆Tair,cond) is 12.6 K and 9.7 K (22.8 °F and 17.5 

°F), respectively, as computed using an energy balance on the airstream, where the condenser 

energy is larger by an amount that reflects the compressor power (in the COP term): 

ɺ )

−1 1 = V ɺ ∆T (2.1) air evap , ( air evap , Qevap ρair Cp air ,

−  1  1 ɺ Qɺ

evap ) 1 

1+  ∆T = V (2.2) air cond , ( air cond ,  COP cool  ρair p air C , 

where the variables represent: 

Cp,air = specific heat of air at constant pressure 

V ɺ Qɺ

evap = condenser airflow per unit evaporator capacity air cond ,

V ɺ Qɺ

evap = evaporator airflow per unit evaporator capacity air evap ,

ρair = density of air 

The air outlet temperatures (Tair,out) are then computed by respectively subtracting or adding the 

∆Tair,evap, ∆Tair,cond, from the Tair,in values. 

The average temperature difference between the refrigerant and the air (∆Tair,ref) for 

Cooling A is specified as 10 K and 3.3 K (18 °F and 6 °F) in the evaporator and condenser, 

respectively, based on experience and measurements at NIST for air-source heat pumps with 

comparable COPs. Table 2-6 shows the average air (Tair,avg) and average refrigerant saturation 

temperatures (Tref,avg) computed using this method. The Cooling B refrigerant saturation 

temperatures are computed using the same ∆Tair and ∆Tair,ref. In heating, the heat exchangers 

operating as the evaporator/condenser swap, so the values of ∆Tair and ∆Tair,ref for the evaporator 

(in cooling) are used for the condenser, and vice versa. The resulting refrigerant saturation 

temperatures are nominally aligned with values observed in heat pump measurements at NIST. 

The HTF inlet temperatures (THTF,in) are adjusted to achieve the target refrigeration saturation 

temperature at a compressor speed of 30 Hz (1800 RPM) with R134a, and then held constant for 

all other speeds. The corresponding HTF outlet temperatures (THTF,out) are defined as: 

T = T + ∆T (2.1) HTF e out , , HTF ,e in , e 

T = T + ∆T (2.2) HTF c out , , HTF ,c in , c 

where ∆Te and ∆Tc are the HTF temperature differences across the evaporator/condenser 

measured by the thermopiles. The THTF,out values differ for each speed; Table 2-6 shows the 

values at 30 Hz (1800 RPM) as a nominal example. As discussed in Section 1, a completely fair 

comparison of refrigerants requires the same HTF inlet and outlet temperatures (McLinden et al., 

1987). However, it is only possible to match either the evaporator or condenser (but not both) 

THTF,out by adjusting the heat pump capacity (via compressor speed); the other THTF,out will be an 

16 
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

uncontrolled value, which will vary somewhat from the R134a value, depending on the COP of 

each fluid (note the HTF mass flow rate is fixed as described in Section 2.3.2.3, so it cannot be 

adjusted to control the other THTF,out). Tests with future refrigerants will be carried out with an 

evaporator or condenser THTF,out equal to the measurements for R134a in cooling or heating, 

respectively, and the other THTF,out will vary for each fluid. 

Table 2-6: Evaporator and condenser fluid temperatures for standard rating tests 

Rating Test 

Tair,in 

°C 

(°F) 

∆Tair 

K 

(°F) 

Tair,out 

°C 

(°F) 

Tair,avg 

°C 

(°F) 

∆Tair,ref 

K 

(°F) 

Tref,avg 

°C 

(°F) 

THTF,in 

°C 

(°F) 

THTF,out 1 

°C 

(°F) 

mHTF ɺ

kg s-1 

lb min-1 

Cooling A 35.0 9.7 44.7 39.9 3.3 43.2 33.9 39.1 0.098 

(95.0) (17.5) (112.5) (103.8) (6.0) (109.8) (93.0) (102.4) (13) 

Condenser Cooling B 27.8 9.7 37.5 32.6 3.3 36.0 24.9 30.9 0.098 

(82.0) (17.5) (99.5) (90.8) (6.0) (96.8) (76.7) (87.6) (13) 

Heating H1 21.1 12.6 33.8 27.4 10.0 37.4 32.0 36.3 0.098 

(70.0) (22.8) (92.8) (81.4) (18.0) (99.4) (89.5) (97.3) (13) 

Cooling A 26.7 

(80.0) 

12.6 

(22.8) 

14.0 

(57.3) 

20.3 

(68.6) 

10.0 

(18.0) 

10.3 

(50.6) 

20.2 

(68.5) 

15.4 

(59.7) 

0.131 

(17.3) 

Evaporator Cooling B 

Heating H1 

26.7 

(80.0) 

12.6 

(22.8) 

14.0 

(57.3) 

20.3 

(68.6) 

10.0 

(18.0) 

10.3 

(50.6) 

19.9 

(67.8) 

14.2 

(57.6) 

0.131 

(17.3) 

8.3 9.7 -1.4 3.5 3.3 0.1 10.3 6.5 0.131 

(47.0) (17.5) (29.5) (38.3) (6.0) (32.3) (50.5) (43.7) (17.3) 

1 Values for R134a at a compressor speed of 30 Hz (1800 RPM) 

2.3.2.2 Operating Parameters 3, 4, & 5: Capacity and Heat Flux 

Baseline capacities were established for R134a at speeds of 23.3 Hz to 36.7 Hz (1400 RPM to 

2200 RPM), in 3.3 Hz (200 RPM) increments, for each Rating Test. Tests with subsequent 

refrigerants will be carried out with compressor speeds that match the capacities from the ɺ baseline tests; this ensures the refrigerants are tested at the same Q A . The target refrigerant- total 

side heat flux range was 2 kW m-2 to 12 kW m-2 (600 Btu h-1 ft-2 to 3800 Btu h-1 ft-2) to coincide 

with values from typical HVAC&R equipment. The average heat flux in the evaporator and 

condenser was controlled by setting the number of active tubes to 10 and 12, respectively. 

Figure 2-6 summarizes the capacity for the baseline tests, where the evaporator capacity varies 

from 1.2 kW to 2.1 kW (0.34 ton to 0.60 ton) and condenser capacity ranges from 1.6 kW to 2.8 

kW (0.45 ton to 0.80 ton). The heat flux in the evaporator varies from 5 kW m-2 to 9 kW m-2 

(1600 Btu h-1 ft-2 to 2900 Btu h-1 ft-2) and in the condenser from 5 kW m-2 to 10 kW m-2 (1700 

Btu h-1 ft-2 to 3200 Btu h-1 ft-2), meeting the target heat flux range. 

17
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heat transfer area = 0.237 m 

 

 

 

 

 

1.4 

1.6 

2.0 

7.0 

7.7 

8.4 

9.1 

2.2 9.4 2.8 

20 25 30 35 40 

Cooling B 

Cooling A 

Heating H1 

Evaporator 

2 

fluid = R134a 

# tubes = 10	�

9.8 

20 25 30 35 40 

Cooling B 

Cooling A 

Heating H1 

Condenser 

heat transfer area = 0.285 m2 

fluid = R134a 

# tubes = 12 

2.6 

8.5 

Average heat flux [kW/m2] 

Capacity [kW]

Average heat flux [W/m2] 

Capacity [kW]

2.4 

2.2 

2.0 

1.8	­ 7.6 

6.8 

1.8 6.0 

1.6 

1.2 5.1 

Compressor frequency [Hz] Compressor frequency [Hz] 

(a) (b) 

Figure 2-6:	­ Capacity and heat flux for baseline R134a tests for (a) evaporator and (b) 

condenser 

2.3.2.3 Operating Parameters 6 & 7: HTF-side Thermal Resistance Ratio 

In typical air-to-air heat pumps the air-side convection provides the majority of the 

resistance to heat transfer. Evaporation heat transfer coefficients are generally larger than those 

for condensation, so the HTF-side resistance ratio (RHTF/Rtotal) is selected in the range of 50 % to 

70 % for the condenser and 70 % to 85 % in the evaporator. The CYCLE_D-HX model reports 

heat exchanger thermal resistances for both the evaporator and condenser when given 

measurements of HTF in/out temperatures, capacity, compressor efficiency, and heat exchanger 

geometry, pressure drop, and log-mean temperature difference. These resistances include the 

heat exchanger total conductance (UA) (the total resistance, Rtotal, is the inverse of UA), the heat 

transfer-side resistance (RHTF) in the evaporator/condenser, and the refrigerant-side thermal 

resistance (Rref) in the evaporator/condenser. Note that the tube wall thermal resistance is 

neglected. 

The MBBHP was operated under varied HTF flowrates ( mɺ HTF ), and the resulting data were 

processed using the CYCLE_D-HX model to determine RHTF/Rtotal for each flowrate. The 

evaporator and condenser HTF flowrates used for the tests are 0.131 kgs-1 and 0.098 kgs-1 (17.3 

lb min-1 and 13 lb min-1) as shown in Table 2-6, which yield RHTF/Rtotal of 80 % and 60 % for the 

baseline Cooling A test with R134a at 30 Hz (1800 RPM). The flowrates are kept constant for 

all other speeds, test conditions, and refrigerants. Maintaining a constant flowrate is important 

for verifying the CYCLE_D-HX model; the model evaluates the HTF-side thermal resistance for 

a reference data set, and then holds this resistance constant as other operating parameters, 

refrigerants, and tube circuit configurations are computationally explored. 

2.3.2.4 Operating Parameters 8 & 9: Subcool and Superheat 

The number of evaporator tubes containing superheated-vapor and condenser tubes with 

subcooled-liquid are controlled to one and two, respectively. It is important to limit the number 

of tubes in superheat and subcooling for comparison with the CYCLE_D-HX model because the 

18
­

5.6 

6.3 
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

model does not consider the pressure drop or area required for heat transfer in these sections, 

rather, it only accounts for transport phenomena in the two-phase section. (Note that the 

superheat and subcooling are included as an input to the model to determine the thermodynamic 

states at the heat exchanger outlets.) The superheat and subcooling are also limited because of 

performance considerations; the single phase fluid exhibits a lower heat transfer coefficient 

compared to a two-phase fluid. The superheat is bound at the low end by the need to prevent 

liquid from entering the compressor. The lower limit for subcooling is governed by a 

requirement of fully subcooled liquid entering the expansion valve; vapor bubbles entering the 

expansion valve cause oscillations in the flow that make it difficult to achieve steady state. In 

general, a superheat range of 3 K to 6 K and a subcooling range of 2 K to 3 K satisfies this 

Operating Parameter requirement. 

Closing the refrigerant needle valve increases the superheat and the subcooling, whereas 

adding refrigerant increases the subcooling and reduces the superheat. The combination of these 

two controls allows for independent regulation of superheat and subcooling. 

2.3.2.5 Operating Parameter 10: LLSL-HX included or bypassed 

The LLSL-HX heat exchanger is included or bypassed using the six valves shown in Figure 

2-1. When the LLSL-HX is bypassed, only one (rather than both) of the valves that stops flow 

through the heat exchanger is closed; this prevents the possibility of trapping subcooled liquid 

between the valves (and presenting the same pressure hazard discussed in Section 2.1. 

19
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3 Data Analysis
­

3.1 Thermodynamic States 

The thermodynamic states numbered 1 through 11 in Figure 2-1 are fixed with two intensive 

properties defined by the measurements and equations presented in Table 3-1. Thermodynamic 

property data for R134a are computed using the Engineering Equation Solver (EES) software 

(Klein, 2015); the software uses the equation of state developed by Tilner-Roth et al. (1994) for 

R134a. Note that property data for mixtures and for fluids not directly available in EES will be 

computed using the NIST REFPROP database (Lemmon et al., 2013). 

Table 3-1: Measurements and equations used to define the thermodynamic states 

State Description Measurement Equations 

1 Compressor inlet, LLSL-HX vapor outlet T1 P1 = P11 – ∆Ps 

2 Compressor cylinder before compression T2 = T1 P2 = P1 

3 Compressor outlet T3 P3 = P4 

4 Condenser inlet T4 P4 = P7 + ∆Pc 

5 Condenser saturated vapor P5 = P4 x5 = 1 

6 Condenser saturated liquid P6 = P7 x6 = 0 

7 Condenser outlet, LLSL-HX liquid inlet T7, P7 

8 Expansion valve inlet, LLSL-HX liquid outlet T8 P8 = P7 

9 Evaporator inlet, Expansion valve outlet P9 h9 = h8 

10 Evaporator saturated vapor P10 = P11 x10 = 1 

11 Evaporator outlet, LLSL-HX vapor inlet T11 P11 = P9 – ∆Pe 

T = temperature, P = pressure, h = enthalpy, x = thermodynamic quality 

3.2 Thermodynamic Performance 

Given the fixed state points identified in Table 3-1 (and all associated intensive properties), 

the key thermodynamic performance metrics are calculated according to the equations in 

Sections 3.2.1 through 3.2.7. Nominal values and uncertainties of the metrics are shown in 

Section 4, and Appendix A, respectively. 

3.2.1 Capacity 

The evaporator and condenser capacities are computed on the refrigerant/HTF-sides 

according to: 

Q ɺ = mɺ (h − h ) (3.1) c ref , ref 4 7 

Q ɺ = mɺ (h − h ) (3.2) e ref , ref 9 11 

Qɺ = mɺ C ∆T −Qɺ (3.3) c, HTF HTF ,c , , p HTF c c ins c , 

20
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Qɺ ɺ = mɺ C ∆T −Q (3.4) e HTF , HTF e , , , p HTF e e ins e ,

where the variables represent: 

mɺ ref = mass flow of refrigerant 

mɺ HTF ,e mɺ HTF ,c = mass flow of HTF in evaporator/condenser 

C C = specific heats (average) of HTF in evaporator/condenser p,HTF,e p,HTF,c 

Qɺ

ins e, Qɺ , = insulation heat leak in evaporator/condenser ins c

∆Te ∆Tc = temperature change of HTF in evaporator/condenser 

The insulation heat leak is a non-trivial 30 W to 50 W, and therefore must be accounted for in the 

heat exchanger capacity. The heat leak is debited from the HTF-side capacity because the HTF 

is the annulus, and therefore thermally interacts with the surrounding ambient air. The heat 

exchangers are divided into active and inactive sections (i.e., tubes where the HTF flows, but the 

refrigerant does and does not, respectively) to compute the heat leak according to: 

 NT NT  e active , e, inactive Q ɺ =UA  (T −T ) + (T −T ) (3.5) ins e , ins e , HTF ,e active avg , , amb e , HTF ,e inactive avg , , amb e ,  NT NT   e total , e, total 

 NT NT  c active , c inactive , Qɺ , =UA ins c , (THTF ,c active avg , , −Tamb c , ) + (T , , , −Tamb c , ) (3.6) ins c  HTF c inactive avg  NT NT   c total , c total , 

where: 

UAins,e UAins,c = insulation conductance for evaporator/condenser 

NTe,total NTe,total = number of tubes in evaporator/condenser total (20) 

NTe,active NTc,active = number of active tubes in evaporator/condenser 

NTe,inactive NTe,inactive = number of inactive tubes in evaporator/condenser 

THTF,e,active,avg THTF,c,active,avg = avg. HTF temp. in the active evaporator/condenser tubes 

THTF,e,inactive,avg THTF,c,inactive,avg = avg. HTF temp. in the inactive evaporator/condenser tubes 

Tamb,e Tamb,c = temp. of ambient air surrounding evaporator/condenser 

The HTF temperatures are taken from the surface-mounted thermocouples, and the ambient air 

temperature is taken from a thermocouple near each of the heat exchangers. The heat exchangers 

are split into active/inactive sections because the slope of the HTF temperature profile is 

different in each section; the heat transfer is only governed by the average temperature difference 

if the slopes (and specific heats) are constant throughout the section. The UAins,e and UAins,c 

values for all 20 tubes are 5.48 W K-1 ±0.31 W K-1 and 6.52 W K-1 ±0.41 WK-1 (10.4 Btu h-1 °F-1 

±0.59 Btu h-1 °F-1 and 12.4 Btu h-1 °F-1 ±0.78 Btu h-1 °F-1). The method used to compute the 

conductance values and their uncertainties is described in Appendix A.8. 

3.2.2 Compressor Power 

The compressor power is computed using the change in enthalpy across the compressor: 

21
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W ɺ = ɺ h comp mref ( 3 − h1 ) (3.7) 

The shaft power is computed for comparison with the compressor power: 

W ɺ = 2π τ N (3.8) shaft

where the variables represent: 

N = compressor shaft speed 

τ = compressor shaft torque 

3.2.3 Coefficient of Performance 

The cooling and heating COP are computed using the refrigerant-side capacity and the 

compressor power applied to the refrigerant: 

COP ɺ = Qɺ Wcomp (3.9) cool e ref ,

ɺ COPheat = Qɺ , Wcomp (3.10) c ref 

3.2.4 Volumetric Capacity 

The volumetric cooling and heating capacity (VCC and VHC) are: 

(mɺ ref VCC = Q v ) (3.11) ɺ

e ref , 1 

ɺ (mɺ ref VHC = Q v ) (3.12) c ref , 1 

where: 

v1 = compressor suction specific volume 

3.2.5 Compressor Efficiency 

The compressor isentropic efficiency is calculated as:
­

h(P , s ) − h 3 1 1 ηs = (3.13) h 3 − h1 

where: 

h(P3,s1) = compressor discharge enthalpy under isentropic compression 

The volumetric efficiency is quantified by the refrigerant displacement normalized by the 

compressor swept volume displacement: 

mɺ v ref 1

ηv = (3.14) D N comp 

where: 

22
­
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

Dcomp = compressor displacement (listed in Table 2-1) 

3.2.6 Superheat and Subcooling 

The evaporator superheat is: 

SH = T −T (P , x =1)	 (3.15) 11 11

and the condenser subcooling is: 

SC = T (P7 , x = 0) −T7	­ (3.16) 

where: 

T(P11, x = 1) = saturated vapor temperature at the evaporator exit pressure
­

T(P7, x = 0) = saturated liquid temperature at the condenser exit pressure
­

The number of refrigerant tubes containing superheated/subcooled fluid in the 

evaporator/condenser is controlled to one and two, respectively, during experimental tests. 

Visual inspection of the temperature profile, such as the one shown in Figure 3-1, is used to 

determine the locations of the various phase transitions. The tube number on the x-axis reflects 

the convention established in Figure 2-5. 

(a)	­ (b) 

Figure 3-1:	­ Temperature profile in (a) evaporator, showing superheat and (b) 

condenser, showing subcooling 

*The location 21 is the exit of the 20th tube 

3.2.7 LLSL-HX Effectiveness 

The LLSL-HX effectiveness is computed as the ratio of the heat transferred on the vapor 

23
­
 

 

 

 

 

side ( Qɺ ) to the maximum possible heat transfer: LLSL v,

Qɺ C (T −T ) LLSL v , p avg , ,1,11 11 1

ε = = (3.17) ɺ LLSL Q LLSL max , MIN( C p avg , ,1,11 , ,7,8 7 ,C p avg )(T −T 11 )

where: 

Cp,avg,1,11 = average specific heats of the refrigerant on the vapor side
­

Cp,avg,7,8 = average specific heats of the refrigerant on the liquid side
­

The heat transfer on the vapor side, rather than liquid side, is chosen for the numerator because 

the larger temperature difference exhibited on the vapor side (due to lower specific heat) results 

in a smaller uncertainty in the effectiveness. 

24
­
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

4 Validation and Baseline Tests with R134a
­

An initial set of tests were carried out with R134a to provide verification of the 

instrumentation and show the nominal uncertainties of the measurements. One of the most 

important validations is to show an energy balance between the refrigerant and the HTF in the ɺ ɺ heat exchangers. The energy imbalance, defined as Qɺ

ref (Qref − QHTF ), for the initial tests is 

shown in Figure 4-1; the imbalance is generally around 1 % to 1.5 % or better in both the 

evaporator and condenser. The imbalance falls well within its uncertainty at the 95 % 

confidence interval, indicating that the estimation of the uncertainty in HTF capacity (3 %) is 

likely too conservative; note that the HTF capacity accounts for more than 90 % of the relative 

uncertainty in the imbalance metric. 

Figure 4-1: Energy imbalance in the evaporator and condenser 

The performance of R134a in the MBBHP, as described by the metrics defined in Section 

3.2, are presented in Figure 4-2 through Figure 4-4. Figure 4-2 (a) and (b) show nominal 

pressure profiles on the low/high pressure sides of the cycle, where the error bars represent 

measurement uncertainty listed in Table 2-3. The compressor power measured both by the shaft 

power and change of enthalpy of the working fluid are shown in Figure 4-3(a), where the 

conversion from mechanical to fluid power is about 65 % efficient. The error bars represent the 

uncertainty calculated by propagating measurement uncertainty through Eqs. (3.7) and (3.8). 

The mass flow ranges from 8 gs-1 to 13 gs-1 (1.1 lb min-1 to 1.7 lb min-1). Figure 4-4 shows the 

COP ranges from 2.4 to 4, and the capacity ranges from 1.6 kW to 2.1 kW (0.45 ton to 0.60 ton). 

The figure shows the uncertainties of the COP and capacity, where the computation and 

tabulation of uncertainty is described further in Appendix A.10 and A.11. The volumetric 

capacity ranges from 2100 kJ m-3 to 2800 kJ m-3 (56 Btu ft-3 to 75 Btu ft-3), and the compressor 

isentropic and volumetric efficiencies are nominally 0.46 and 0.55; uncertainties in all three of 

these metrics are shown in the figures and are computed using a similar method as the 

capacity/COP, but the computation is not discussed in detail in this report. 

Figure 4-3 (b) also shows the mass flux, which is used to determine the two-phase flow 

regimes indicated in Figure 4-5. The refrigerant generally entered the annular flow regime in the 

evaporator and condenser, though for some tests with low mass flux the refrigerant never left the 

stratified-wavy regime. 

25
­
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

1100 

1150 

1200 450 

C

o

m

pre

s

s

or p

o

w

er [kW]	­ Pressure [kPa] 

# evaporator tubes = 10 

# condenser tubes = 12 

fluid = R134a 

LLSL-HX = no 

condenser outlet (P7) 

rating test = cooling A 

condenser inlet (P4) 

Pressure [kPa] 

20 25 30 35 40 

Compressor frequency [Hz] 

# evaporator tubes = 10 

# condenser tubes = 12 

fluid = R134a 

LLSL-HX = no 

evaporator outlet (P11) 

compressor suction (P1) 

rating test = cooling A 

evaporator inlet (P9) 

400 

350 

1050 300 

20 25 30 35 40 

Compressor frequency [Hz] 

(a)	­ (b) 

Figure 4-2:	­ Heat pump cycle pressures on the (a) low pressure side and (b) high
­

pressure side
­

1.2 0.013 

20 25 30 35 40 

Compressor frequency [Hz] 

# evaporator tubes = 10 

# condenser tubes = 12 

fluid = R134a 

LLSL-HX = no 

cooling A 

cooling B 

heating H1 

pow er to w orking fluid (Wcomp) 

shaft pow er (Wshaf t) 

Mass flow [kg/s]

20 25 30 35 40 

# evaporator tubes = 10 

# condenser tubes = 12 

fluid = R134a 

LLSL-HX = no 

cooling A 

cooling B 

heating H1 

0.012 

Mass flux [kg/s-m2] 

0.9 200 

0.011 

180 0.01 

0.6 

0.009 

0.008 

0.3 

Compressor frequency [Hz] 

(a)	­ (b) 

Figure 4-3:	­ Performance metrics including (a) compressor power and (b) mass flow and 

mass flux 

26
­

140 

160 

220 
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

V

olumetric capacity [kJ/m3] COP 

Volumetric efficiency (

ηv) 

(a) (b) 

20 25 30 35 40 

2 

2.5 

3 

3.5 

4 

4.5 

Compressor frequency [Hz] 

# evaporator tubes = 10 

# condenser tubes = 12 

cooling A 

cooling B 

heating H1 

fluid = R134a 

LLSL-HX = no 

20 25 30 35 40 

1.5 

1.6 

1.7 

1.8 

1.9 

2 

2.1 

Compressor frequency [Hz] 

C

a

p

a

city [kW] 

cooling A 

cooling B 

heating H1 

# evaporator tubes = 10 

# condenser tubes = 12 

fluid = R134a 

LLSL-HX = no 

2800 0.7 0.7 

# evaporator tubes = 10 

# condenser tubes = 12 

fluid = R134a 

LLSL-HX = no 

cooling A 

cooling B 

heating H1 

20 25 30 35 40 

0.4 

0.5 

0.6 

0.3 

0.4 

0.5 

0.6 

# evaporator tubes = 10 

# condenser tubes = 12 

fluid = R134a 

LLSL-HX = no 

cooling A 

cooling B 

heating H1 

cooling B 

cooling A 

heating H1 

2600 

2400 

2200 

Isentro

pic efficiency ( ηs) 

20 

2000 

25 30 

Compressor frequency 

35 

[Hz] 

40 

Compressor frequency [Hz] 

(c) (d) 

Figure 4-4: Performance metrics including (a) COP, (b) capacity, (c) volumetric 

capacity, and (d) compressor isentropic and volumetric efficiency 

27
­
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

50 

Flow boiling with microfin tube	­ Flow condensing with microfin tube 350 350 

300 300 

S 

heat flux 7.0 kW/m2 

= 

= A 

I 

SG 

S+SW 

D 

Tsat = 8 C 

fluid = R134a 

flow range during tests 

M 

SW 

= 

Tsat 43 C 

fluid = R134a 

heat flux 7.7 kW/m2 

flow range during tests 

A 

I 

SG 

S+SW

SW 

0 

50 

S 

Mass Flux [kg/s-m2] 250 

200 

150 

100 

Mass Flux [kg/s-m2] 250 

200 

150 

100 

0 

0	­ 0.2 0.4 0.6 0.8 1 0 0.2 0.4 0.6 0.8 1 

Vapor quality Vapor quality 

A = Annular D = Dryout I = Intermittant M = Mist A = Annular D = Dryout I = Intermittant M = Mist 

S = Stratified SG = Slug SW = Stratified-Wavy S = Stratified SG = Slug SW = Stratified-Wavy 

(a)1,3	­ (b)2,3 

Figure 4-5:	­ Flow regime map for microfin tube (a) evaporation and (b) condensation 

1 Flow boiling regime map from (Wojtan et al., 2005)
­

2 Flow condensation regime map from (Hajal et al., 2003)
­

3 Boundary between Intermittent/Annular regimes from (Thome, 2007), using equation 12.6.1 with a coefficient
­

of 0.678, for helical microfins 

5 Conclusions and Future Work 

This report shows that the test apparatus is capable of measuring the performance of 

low-GWP refrigerants with high fidelity, including less than 0.5 % uncertainty in both capacity 

and COP. The energy balance closure on the evaporator and condenser was better than 1.5 % 

and will be continually tracked to ensure continued high quality data are collected. 

Future tests will include a second baseline refrigerant R410A, followed by a series of 

comparative tests with low-GWP refrigerants and refrigerant mixtures. The HTF flowrates will 

remain fixed from the R134a baseline tests, so that the CYCLE_D-HX simulations of the test 

data can be carried out for all test data based on a single reference data set. After the model 

results are verified, the model will be used to identify additional candidate refrigerants to test in 

the MBBHP test apparatus. 

28
­
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

6 References
­

AHRI. (2008). 2008 Standard for Performance Rating of Unitary Air-Conditioning & AirSource Heat Pump Equipment. Air-Conditioning, Heating, and Refrigeration Institute, 

Arlington, VA, United States. Retrieved from www.ahrinet.org 

AHRI. (2015). Participants' Handbook: AHRI Low-GWP Alternative Refrigerants Evaluation 

Program (Low-GWP AREP). Arlington, VA, United States: Air-Conditioning, Heating, 

and Refrigeration Institute. Retrieved from 

http://www.ahrinet.org/site/514/Resources/Research/AHRI-Low-GWP-AlternativeRefrigerants-Evaluation 

Brown, J. S., Domanski, P. A., & Lemmon, E. (2011). Standard Reference Database 49, 

CYCLE_D: NIST Vapor Compression Cycle Design Program, Version 5.0. Retrieved 

from http://www.nist.gov/srd/nist49.cfm 

Domanski, P. A., Brown, S. J., Heo, J., Wojtusiak, J., & McLinden, M. O. (2014). A 

thermodynamic analysis of refrigerants: Performance limits of the vapor compression 

cycle. International Journal of Refrigeration, 38, 71-79. 

doi:10.1016/j.ijrefrig.2013.09.036 

EPA. (2015). 2015 North American Amendment Proposal to Address HFCs under the Montreal 

Protocol. Retrieved from http://www.epa.gov/ozone/intpol/mpagreement.html and 

http://www.epa.gov/ozone/intpol/HFC_Amendment_2015_Summary.pdf 

EU. (2014, April 16). Regulation (EU) No 517/2014 of the European Parliament and of the 

Council of 16 April 2014 on fluorinated greenhouse gases and repealing Regulation (EC) 

No 842/2006. Journal of the European Union. Retrieved from http://eurlex.europa.eu/legal-content/EN/TXT/?uri=uriserv:OJ.L_.2014.150.01.0195.01.ENG 

Fortin, T. J. (2015). Personal communication. 

Hajal, J. E., Thome, J. R., & Cavallini, A. (2003). Condensation in horizontal tubes, part 1: twophase flow pattern map. International Journal of Heat and Mass Transfer, 3349-3363. 

doi:10.1016/S0017-9310(03)00139-X 

Harr, L., Gallagher, J. S., & Kell, G. S. (1984). NBS/NRC Steam Tables. Hemisphere Publishing 

Co. 

Kazakov, A., McLinden, M. O., & Frenkel, M. (2012). Computational Design of New 

Refrigerant Fluids Based on Environmental, Safety, and Thermodynamic Characteristics. 

Industrial & Engineering Chemistry Research, 51, 12537-12548. doi:10.1021/ie3016126 

Kim, Y., Payne, V., Choi, J., & Domanski, P. A. (2005). Mass flow rate of R-410A through short 

tubes working near the critical point. International Journal of Refrigeration, 28, 547-553. 

doi:10.1016/j.ijrefrig.2004.10.007 

Klein, S. A. (2015). Engineering Equation Solver, v 9.908. Retrieved from 

http://www.fchart.com/ees/ 

Lemmon, E. W., Huber, M. L., & McLinden, M. O. (2013). NIST Standard Reference Database 

23, Reference Fluid Thermodynamic and Transport Properties (REFPROP), Version 9.1. 

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National Institute of Standards and Technology. Retrieved from
­

http://www.nist.gov/srd/nist23.cfm
­

McLinden, M. O., & Radermacher., R. (1987). Methods for Comparing the Performance of Pure 

and Mixed Refrigerants in the Vapour Compression Cycle. International Journal of 

Refrigeration, 10(6), 318-325. doi:10.1016/0140-7007(87)90117-4 

McLinden, M. O., Kazakov, A., Heo, J., Brown, J. S., & Domanski, P. A. (2014). A 

thermodynamic analysis of refrigerants: Possibilities and tradeoffs for Low-GWP 

refrigerants. International Journal of Refrigeration, 38, 80-92. 

doi:10.1016/j.ijrefrig.2013.09.032 

Pannock, J., & Didion, D. A. (1991). The Performance of Chlorine-Free Binary Zeotropic 

Refrigerant Mixtures in a Heat Pump, NISTIR 4748. Internal Report, National Institute of 

Standards and Technology, U.S. Department of Commerce, Gaithersburg. Retrieved from 

http://fire.nist.gov/bfrlpubs/build91/art002.html 

Solomon, S., Qin, D., Manning, M., Chen, Z., Marquis, M., Averyt, K. B., . . . Miller, H. L. 

(2007). Contribution of Working Group I to the Fourth Assessment Report of the 

Intergovernmental Panel on Climate Change. New York, NY, USA: Cambridge 

University Press. Retrieved July 29, 2015, from 

http://www.ipcc.ch/publications_and_data/publications_ipcc_fourth_assessment_report_ 

wg1_report_the_physical_science_basis.htm 

Taylor, B. N., & Kuyatt, C. E. (1994). Guidelines for Evaluating and Expressing the Uncertainty 

of NIST Measurement Results, National Institute of Standards and Technology Technical 

Note 1297. 

Thome, J. R. (Updated in 2007). Wolverine Heat Transfer Engineering Data book III. Germany: 

Weiland-Werke AG. Retrieved from http://www.wlv.com/heat-transfer-databook/ 

Tilner-Roth, R., & Baehr, H. D. (1994). An International Standard Formulation for the 

Thermodynamic Properties of 1,1,1,2-Tetrafluoroethane (HFC-134a) for Temperatures 

from 170 K to 455 K and Pressures up to 70 MPa. J. Phys. Chem, Ref. Data, 23(5). 

Wojtan, L., Ursenbacher, T., & Thome, J. R. (2005). Investigation of flow boiling in horizontal 

tubes: Part I—A new diabatic two-phase flow pattern map. International Journal of Heat 

and Mass Transfer, 2955-2969. doi:10.1016/j.ijheatmasstransfer.2004.12.012 

30
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Appendix A: Uncertainty Analysis
­

A.1 Symbols Used in Uncertainty Analysis
­

Symbol Units Definition 

k -- Expanded uncertainty coverage factor 

n -- Number of data points 

Qɺ kW Energy transfer 

SUA,ins,e W K -1 Standard deviation of evaporator insulation conductance 

SX -- Standard deviation of measured quantity “X” 

T °C Temperature 

T1 °C Temperature at cold end of thermopile 

Tcal °C Temperature of RTD used for calibrations 

TTC,cal °C Calibrated thermocouple measurement 

TTC,CJC °C Thermocouple measurement with Cold Junction Compensation 

∆Tlm K Log-mean temperature difference 

∆TTC,cal K Thermocouple calibration temperature difference 

∆TTP K Temperature difference of thermopile 

UA W K -1 Thermal conductance 

U UA ins e , , W K-1 Uncertainty of average of evaporator insulation conductance 

UTC °C Total instrument uncertainty of thermocouple 

UTC,cal °C Thermocouple instrument uncertainty from calibrating device uncertainty 

UTC,fit °C Instrument uncertainty of thermocouple due to fit of calibration regression 

UX -- Uncertainty of measured quantity “X” 

U X -- Uncertainty of average of measured quantity “X” 

UY Uncertainty of calculated quantity “Y” 

V V Voltage 

V1 V Thermopile voltage of cold end of thermopile (referenced to ice-point) 

VCC kJ m -3 Volumetric Cooling Capacity 

VHC kJ m -3 Volumetric Heating Capacity 

VTC V Thermocouple voltage 

∆VTP V Thermopile voltage 

X -- Measured quantity 

Y -- Calculated quantity 

31
­
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

Subscript Definition 

amb Ambient air
­

c Condenser
­

e Evaporator
­

HTF Heat Transfer Fluid
­

in Inlet
­

ins Insulation
­

out Outlet
­

1 to 11 Refrigerant thermodynamic states as defined by Figure 2-1
­

Abbreviation Definition 

COP Coefficient of Performance (W capacity per W of input power)
­

CJC Cold Junction Compensation (for thermocouples)
­

DAQ Data Acquisition
­

DSC Differential Scanning Calorimeter (used to measure fluid specific heat)
­

EES Engineering Equation Solver (software used to reduce data)
­

HTF Heat Transfer Fluid
­

MBBHP Mini Breadboard Heat Pump
­

RPM Revolutions per Minute (compressor shaft)
­

RTD Resistance Temperature Detector
­

A.2 General Remarks 

The uncertainty analyses for key performance metrics are presented in this section. All 

uncertainties are estimated based on instrument uncertainties computed at a 95 % confidence 

interval. 

A.3 Thermocouples with CJC Compensation 

The surface-mounted thermocouples, as well as the thermocouples measuring the air 

temperature near the evaporator/condenser, are compensated at the DAQ using a thermistorbased Cold Junction Compensation (CJC). A calibration (∆Tcal) is applied to the resulting 

temperature reading (TTC,CJC) to compute the calibrated temperature (TTC,cal) according to: 

T = T + ∆T (A.1) TC ,cal , TC CJC TC cal , 

where the calibration is fit to a 2nd order polynomial (the order of this curve fit, and others 

presented in Appendix A, were selected based on t-test of polynomial coefficients): 

= T − a T + a T 2 ∆T , cal T , = a0 + 1 TC CJC 2 TC CJC (A.2) TC cal TC CJC , , 

The polynomial coefficients (a0, a1, a2) are fit using the least square error method, and Tcal is the 

temperature measured by the calibrating instrument (two calibrated RTDs read by a precision 

digital thermometer, expanded uncertainty of ±0.02 °C). The thermocouples were calibrated over 

a temperature range of -10 °C to 60 °C, and this calibration was repeated five times over the 

32
­
course of two weeks; Figure A-1 (a) shows the calibration data and polynomial curve fits for a 

representative thermocouple. Each calibration shows a significant offset from the others; this is 

likely a manifestation of the repeatability error of the CJC, which has an overall uncertainty of 

±0.5 °C. Figure A-1 (b) shows the aggregation of the five data sets and the resulting curve fit 

and 95 % confidence interval (±0.6 °C). The uncertainty of the calibrated RTDs is more than an 

order of magnitude smaller than the curve fit confidence interval, and is therefore neglected. The 

CJC thermocouple instrument uncertainty listed in Table 2-3 is therefore ±0.6 °C. 

0.5 0.5 

Temperature calibration (

∆

Tcal) [K] 

Temperature calibration (

∆

Tcal) [K]

0 

-0.5 

-1 

data curve fit 

95% confidence interval 

0 

-0.5 calibration 1 calibration 2 calibration 3 calibration 4

-1 calibration 5 curve fits 

-1.5 -1.5 

-10 0 10 20 30 40 50 60 -10 0 10 20 30 40 50 60 

Temperature (TTC,CJC) [°C] Temperature (TTC,CJC) [°C] 

(a) (b) 

Figure A-1: Temperature calibration for CJC compensated thermocouples where 

(a) data are divided by each of the 5 calibrations and (b) all data are 

combined. 

A.4 Thermocouples with Ice-Water Bath Compensation 

A lower level of uncertainty is achieved for the in-stream probe thermocouples, which 

measure the refrigerant temperatures at key thermodynamic states, by utilizing an ice-water bath 

reference junction. The ice-water bath consists of a vacuum insulated Dewar filled with crushed 

ice and water; the reference junction is submerged in the bath inside of a water-filled test tube. 

Using the ice-water bath removes the large uncertainty associated with the electronic CJC on the 

DAQ system. The thermocouples were calibrated against the RTDs (same ones used to calibrate 

thermocouples with CJC), where the resulting data and curve fit are shown for a representative 

thermocouple in Figure A-2. The data were fit to 5th order polynomials according to: 

5 

T aV i (A.3) TC cal , = ∑ i TC 

i=0 

where VTC is the thermocouple voltage. The resulting instrument uncertainty (±0.08 °C) listed in 

Table 2-3 is the combination of the curve fit standard error (UTC,fit, k = 2) and the calibrating 

RTD error (UT,cal): 

2 2 2 2 U = ± U +U = ± 0.079 + 0.02 = ±0.08 °C (A.4) TC TC, fit T ,cal 

33 
Temperature [°C]

80 

data 

60 curve fit

40 

20 

0 

-20 

-0.0005 0 0.0005 0.001 0.0015 0.002 0.0025 

Voltage [V] 

Figure A-2: Ice-water bath referenced thermocouple calibration data 

A.5 Thermopiles in Heat Transfer Fluid 

The 13-junction thermopiles were calibrated by immersing one end in an ice-water bath 

(inside a water-filled test tube) and the other end in a temperature controlled bath. The 

temperature in the controlled bath was measured using the two calibrated RTDs. A thermopile 

voltage-temperature dataset was generated and fit to a 4th order polynomial. Figure A-3 shows 

the calibration data and the curve fit, and Table A-1 shows the polynomial coefficients. The 

calibration data are fit to the equation: 

N =4 

i T = ∑aTP i, V (A.5) 

i=0 

The temperature difference for a thermopile is computed according to: 

∆T = T V ( + ∆V ) −T (A.6) TP 1 TP 1 

where ∆VTP is the measured thermopile voltage and T1 is the temperature measured at the cold 

end of the thermopile by a CJC-corrected thermocouple. Eq. (A.5) is substituted into Eq. (A.6) 

to calculate the thermopile temperature difference: 

N =4 

∆TTP = ∑aTP i, (V1 + ∆VTP ) −T1 (A.7) 

i=0 

where V1 (cold-end voltage) is computed by numerically solving Eq. (A.5) where T is set to (T1): 

N =4  i  V1 = root T1 −∑aTP,i V1 = 0 (A.8)  i=0 

The thermopile uncertainties are computed with respect to the curve fit, voltage 

measurement, and temperature measurement at the cold end of the thermopile (characterized 

respectively by the coefficient uncertainties in Table A-1, the thermopile voltage uncertainty in 

Table 2-3, and the CJC corrected thermocouple in Table 2-3). The EES (Klein, 2015) software 

34 
is used to numerically compute the partial derivatives and subsequent thermopile uncertainties. 

Figure A-4(a) shows the contours of uncertainty in the temperature difference measured by the 

evaporator/condenser HTF thermopiles. 

Figure A-4(b) shows the contours of relative contribution to uncertainty in the thermopile 

measurement. Each of the uncertainty sources is dominant in different regions: 

• voltage measurement: dominant at low temperature differences (less than 1 K) 

• curve fit error: dominant at high cold end temperatures (greater than 50 °C) 

• cold end temperature: dominant at moderate cold end temperatures -5 °C to 20 °C 

and temperature differences 2.5 K to 20 K 

As noted in this discussion, the thermopile uncertainty is a complex function of the curve fit, 

voltage measurement, and cold end temperature measurement. For the temperature difference 

range expected for this application, 1 K to 10 K (1.8 °F to 18 °F), the upper bound of uncertainty 

is about 0.015 K (0.027 °F); for simplicity, this is the value listed in Table 2-3. Note that the 

contribution to the uncertainty by the calibrating instrument (two calibrated RTDs) is negligible 

here as the error in slope (which is the error that propagates into the thermopile temperature 

difference) is less than 0.000248 K per K of temperature difference (e.g. 0.00248 K for a 10 K 

temperature difference). 

Table A-1: Condenser & evaporator 60 

thermopile polynomial coefficients 

50 Uncertainty 

40 Coeff. Value (k=2), 95%

Tem

perature [°C] 

aTP,e,0 3.9243 E-02 ±4.1 E-04 30 

aTP,e,1 2.0243 E+03 ±1.2 E-01 

20 

aTP,e,2 -5.5466 E+03 ±2.2 E+01 TP1 10 TP2 aTP,e,3 3.9023 E+04 ±1.3 E+03 TP1 (fit) 

TP2 (fit) aTP,e,4 -2.6229 E+05 ±2.2 E+04 

 

0 

-10 

3.8494 E-02 ±5.3 E-04 -0.01 0 0.01 0.02 0.03 0.04 aTP,c,0

Voltage [V]
­ aTP,c,1 2.0241 E+03 ±1.5 E-01 

Figure A-3: Thermopile calibration data aTP,c,2 -5.5190 E+03 ±2.8 E+01 

aTP,c,3 3.8551 E+04 ±1.6 E+03 

aTP,c,4 -2.5440 E+05 ±2.9 E+04 

35
 

 

 

 

 

 

 

 

 

 

 

 

 

 

 

-5 

0 

10 

20 contours of thermopile uncertainty [K] contours of contribution to uncertainty [%] 20 

*balance is T1 0.04 

0.035 

Temperature difference (

∆

TTP) [K] 

15 0.03 

0.025 

0.02 

0.015 

0.01 

5 0.001 

0.002 0.005 

0.004 

0.003 

Temperature difference (

∆

TTP) [K] 

15 

10 

curve fit* 

80 60 40 20 

5 voltage (∆VTP)* 

20 

40 

0 

60 80 

0 10 20 30 40 50 

Temperature at cold end (T1) [°C] 

60 

-5 

0 10 20 30 40 50 

Temperature at cold end (T1) [°C] 

60 

(a) (b) 

Figure A-4: Contours of thermopile (a) uncertainty and (b) relative contribution to 

uncertainty 

A.6 Evaporator Heat Transfer Fluid Specific Heat 

The evaporator HTF is a potassium formate and water heat transfer fluid that is freeze 

protected to -40 °C. The manufacturer’s specific heat data were found to have error of at least 

10 %, so a sample of the fluid was analyzed in a Differential Scanning Calorimeter (DSC) by the 

NIST-Boulder Applied Chemicals and Materials Division. The temperature dependent specific 

heat and the 2nd order curve fit to the data are shown in Figure A-5; the curve fit coefficients and 

their uncertainties are listed in Table A-2. Figure A-5 also shows the manufacturer’s reported 

specific heat data for comparison. The exact uncertainty of the DSC measurement was not 

available at the time of writing, but was estimated to be about 2 % to 3 % of reading based on the 

experience of the DSC operator (Fortin, 2015). An uncertainty of 3 % of reading is therefore 

assigned as a conservative estimate. Both the curve fit and DSC calibration uncertainties are 

included in the calculation of overall uncertainties. 

36
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DSC measurements

curve fit

3 Table A-2: Evaporator HTF specific 

heat polynomial coefficients 

2.9 Uncertainty 

DSC measurements 

curve fit 

manufacturer's data 

-10 0 10 20 30 

Specific heat (

Cp) [J/g-K] 

Coeff. Value (k=2), 95% 

2.8 aCp,e,0 2.5643 E+00 ±1.6 E-03 

aCp,e,1 1.1015 E-03 ±2.1 E-04 

2.7 aCp,e,2 1.7422 E-05 ±9.5 E-06 

2.6 

2.5 

Temperature [°C] 

Figure A-5: Specific heat measurements for 

evaporator HTF 

A.7 Condenser Heat Transfer Fluid Specific Heat 

The condenser HTF is water; the specific heat data are computed in EES using the data from 

(Harr et al., 1984) and are shown in Figure A-6. A conservative estimate of ±3 % is used for the 

uncertainty in the specific heat. 

4.23 

0 20 40 60 80 100 

Specific heat (

Cp) [kJ/kg-K] 4.22 

4.21 

4.2 

4.19 

4.18 

Temperature [°C] 

Figure A-6: Specific heat of condenser HTF (water) 

A.8 Evaporator and Condenser Insulation Conductance 

The total evaporator and condenser insulation conductances (UAins,e and UAins,c) for thermal 

interaction with the ambient air were measured during an initial test where the HTFs were 

circulated at various flow rates through the apparatus at an elevated temperature, while the 

refrigerant tubes were under vacuum. The temperature change of the HTF across the heat 

37
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exchanger was assumed to be entirely attributed to insulation heat leak, and therefore an 

effective insulation conductance could be computed. The resulting UAins,e and UAins,c values for 

all 20 tubes are shown in Figure A-7. Average values of conductances for the evaporator and 

condenser were calculated as 5.48 W K-1 ±0.31 W K-1 and 6.52 W K-1 ±0.41 W K-1 

(10.4 Btu h-1 °F-1 ±0.59 Btu h-1 °F-1 and 12.4 Btu h-1 °F-1 ±0.78 Btu h-1 °F-1), respectively, where the 

uncertainty accounts for both instrumentation error and test variance. The evaporator insulation 

conductance for these tests is computed according to: 

Qɺ

ins e , mɺ HTF e , C , , (THTF in e , , −T p HTF e HTF out e , , ) UA = = (A.9) ins e , ∆T T −T − T −T lm HTF , ,e ( HTF in e , , amb e , ) ( HTF ,out, e amb e , )

 (T −T )  HTF in e , , amb e , ln   (T , , −Tamb e ) 

  HTF out e , 

where the variables represent: 

Cp HTF , ,e = evaporator HTF specific heat 

mɺ HTF ,e = evaporator HTF mass flow 

Qɺ = evaporator insulation heat leak ins e,

Tamb,e = ambient air temperature surrounding evaporator 

THTF,in,e = evaporator HTF inlet temperature 

THTF,out,e = evaporator HTF outlet temperature 

∆Tlm,HTF,e = evaporator log-mean temperature difference 

and the uncertainty of the average value is calculated according to: 

U = k S , , n (A.10) UA ins e , , UA ins e 

where the variables represent: 

k = expanded uncertainty coverage factor (k = 2) for 95 % confidence interval 

SUA,ins,e = standard deviation of conductances 

n = number of conductance data points 

The condenser conductance is computed using the same method with the corresponding 

measurements from that heat exchanger. 

38
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4 

5 

6 

7 

8 

Insulation conductance (UAins) [W/K]

UAins,cond 

UAins,ev ap 

average for evaporator 

average for condenser 

0 0.05 0.1 0.15 

Mass flow [kg/s] 

Figure A-7: Insulation conductance for evaporator and condenser 

A.9 Uncertainty Analysis Software 

The uncertainty analyses on each of the performance metrics shown in Figure 4-1 through 

Figure 4-4 were calculated using the uncertainty propagation capability in EES. The software 

numerically computes the performance metric uncertainty by evaluating the partial derivatives 

with respect to each relevant measurement and applying the specified measurement uncertainty. 

Assuming the measurements are independent and uncorrelated, the uncertainty of a quantity is 

computed according to (Taylor & Kuyatt, 1994): 

 ∂Y 2 

2 

UY = ∑ UX  (A.11) 

i  ∂Xi 

i 

where the variables represent: 

Y = calculated quantity 

Uy = uncertainty of calculated quantity 

Xi = measurement 

UXi = uncertainty of measurement 

39
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A.10 Moving Window and Steady State Uncertainty 

A moving window average was used to compute the measurements at steady state during the 

tests. In order to determine a proper window size, the variation of measurement steady state 

uncertainty with window size was examined. The steady state uncertainty is defined as the 

expanded standard uncertainty for the average reading within the window: 

S X U = k (A.12) X n 

where: 

k = coverage factor (k = 2) for 95 % confidence interval 

n = number of samples in the window 

Sx = standard deviation of measurement 

X = average value of measurement 

The data for analysis were generated by holding the MBBHP at a nominal steady operating 

condition and recording the measurements at 1-minute intervals (same as collection rate for all 

performance testing) for an extended time (3 hours). Figure A-8 shows the fluctuation of steady 

state uncertainty for the evaporator inlet and outlet temperatures (T9 and T11). The figure 

highlights two characteristic trends observed for all of the instruments, both of which suggest the 

need for a large moving window: (1) the steady state uncertainty decreases as the window size 

increases, and (2) the steady state uncertainty is highly oscillatory for small windows (the 

oscillations are undesirable because short-term fluctuations can have a disproportionate impact 

on the averaged result). A 30 sample moving window was selected to achieve low uncertainty 

while allowing for collection of data in a timely manner. 

Figure A-8 shows that the measurement steady state uncertainty is not constant in time (it 

would be constant if the fluctuations at steady state were purely random), but rather exhibits 

drifting and periodic movements. This is possibly due to hunting by the PID controllers 

regulating HTF fluid temperatures, changes in air temperature in the lab space, or some other 

uncontrolled and/or uncharacterized phenomena. These non-random fluctuations are accounted 

for by using the maximum steady state uncertainty value observed during the extended steady 

state test. For example, Figure A-8(a) shows the maximum steady state uncertainty for the 

evaporator outlet (T9) is ±0.06 °C; this value and values for all other sensors are listed in Table 

2-3. The data in the moving window are only recorded when steady state uncertainties are below 

these established thresholds. The total uncertainty (±0.09 °C), also shown in Table 2-3, is 

subsequently computed by adding the instrument uncertainty (±0.08 °C) and steady state 

uncertainty (±0.06 °C) in quadrature. 

Note that different sensors of the same type exhibit different levels of fluctuation; as a 

conservative estimate, the largest uncertainty of any one is applied to all of them. For example, 

Figure A-8(b) shows the fluctuations of steady state uncertainty for the evaporator inlet 

temperature (T9) are much smaller than for the evaporator outlet (T11). In fact, the evaporator 

outlet temperature (T11) shows the most fluctuation of all the ice-water bath compensated 

thermocouple probes; therefore, its value (±0.06 °C) is used in Table 2-3 to represent them all. 

40
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0.5 0.5 

0.1 0.1 T,s [C] 

0.01 0.01 

U teady 

0.002 

(a) 

Time elapsed [s] 

0.002 

(b) 

Time elapsed [s] 

Figure A-8:	­ Uncertainty in temperature due to fluctuations at steady state for (a) 

evaporator outlet and (b) evaporator inlet 

A.11 Uncertainty Results Summary 

The detailed instrument uncertainty analysis for performance metrics including the 

refrigerant/HTF-side capacities and the COPs are tabulated in Appendices A.11.1 (cooling) and 

A.11.2 (heating). The tables show the performance metric and its total uncertainty (including 

both instrument and steady state contributions), as well as the measurements that contribute to 

the uncertainty. The measurement values, uncertainties, partial derivatives (of performance 

metric with respect to the measurement), and relative errors (contribution to the overall error in 

the performance metric) are also shown, where the measurements are listed in order of 

decreasing relative error. Uncertainties for other performance metrics (e.g. VCC/VHC, ηisen) 

were analyzed using similar techniques, but are not shown in detail in this report. Note that 

Appendices A.11.1 and A.11.2 show sample calculations for a single nominal operating 

condition, whereas the uncertainty values presented in the Section 4 figures are computed for 

each test run. 

The uncertainty in cooling performance metrics are shown for a nominal operating point for 

R134a during a Cooling A test with a compressor speed of about 30 Hz (1800 RPM). Similarly, 

the uncertainty in heating performance metrics are shown for R134a during a Heating H1 test 

with a compressor speed of about 30 Hz (1800 RPM). 

A.11.1 Uncertainty Results Summary: Cooling 

Table A-3 shows that the dominant sources of instrument uncertainty for the refrigerant-side 

capacity are refrigerant mass flow (66 %), expansion valve inlet temperature (17 %), evaporator 

exit pressure (8 %), and evaporator exit temperature (7 %). Table A-4 shows the dominant 

source of instrument uncertainty for the HTF-side capacity is the HTF specific heat (97 %, 

estimate for instrument used to measure fluid properties). Table A-5 shows the dominant 

sources of instrument uncertainty for cooling COP are evaporator inlet pressure (39 %), 

compressor discharge temperature (25 %), and compressor suction temperature (18 %). 

8000 12000 16000 20000 

UT,steady [C] 

increasing 

time window 

10 min 

30 min 

60 min 

05 min 

maximum (0.06 °C) 

8000 12000 16000 20000 

increasing 

time window 

10 min 

30 min 

60 min 

05 min 

41
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Table A-3: Uncertainty of refrigerant-side evaporator cooling capacity 

Total Partial Relative 

Parameter Symbol Value Uncertainty Derivative Error 

Refrigerant-side evaporator capacity Qe,ref ɺ 1.692 kW ±0.0035 kW 

(0.20%) -- --

Refrigerant mass flow mref ɺ 11.00 g s-1 ±0.024 g s-1 1.538 E-01 66 % 

Refrigerant temp. at expansion valve inlet T8 38.16 C ±0.09 °C -1.634 E-02 17 % 

Refrigerant press. at evaporator inlet P9 412 kPa ±3.5 kPa -2.817 E-04 8 % 

Refrigerant temp. at evaporator exit T11 12.09 C ±0.09 °C 1.009 E-02 7 % 

Refrigerant press. at condenser exit P7 1104 kPa ±3.5 kPa 2.103 E-06 0 % 

Refrigerant evaporator differential press. ∆Pe 43.92 kPa ±0.8 kPa 2.817 E-04 0 % 

Table A-4: Uncertainty of HTF-side evaporator cooling capacity 

Total Partial Relative 

Parameter Symbol Value Uncertainty Derivative Error 

±0.05 kW HTF-side evaporator capacity Qɺ 1.672 kW -- -- e,HTF (3.0%) 

Evaporator HTF capacity cp HTF ,e 2.59 J g-1 K-1 ±3 % of cp 6.369 E-01 97 % calibration uncertainty HTF ,e 

Temperature of ambient air 

surrounding evaporator Tamb,e 23.0 °C ±0.6 °C 5.474 E-03 1 % 

HTF mass flow mHTF ,e ɺ 131.2 g s-1 ±0.26 g s-1 1.257 E-02 1 % 

Evaporator HTF thermopile voltage ∆VTP,e 2.512 mV ±0.007 mV 6.537 E-01 1 % 

Evaporator HTF thermopile aTP,e,0 3.9243 E-02 ±4.1 E-04 4.047 E-03 0 % 

polynomial coefficients aTP,e,1 2.0243 E+03 ±1.2 E-01 8.832 E-04 0 % 

aTP,e,2 -5.5466 E+03 ±2.2 E+01 1.559 E-05 0 % 

aTP,e,3 3.9023 E+04 ±1.3 E+03 2.106 E-07 0 % 

aTP,e,4 -2.6229 E+05 ±2.2 E+04 2.554 E-09 0 % 

Evaporator HTF capacity aCp,e,0 2.5643 E+00 ±1.6 E-03 6.369 E-01 0 % 

polynomial coefficients aCp,e,1 1.1015 E-03 ±2.1 E-04 1.135 E+01 0 % 

aCp,e,2 1.7422 E-05 ±9.5 E-06 2.024 E+02 0 % 

Temperature at cold end of 

evaporator HTF thermopile THTF,e,out 15.42 °C ±0.6 °C -2.963 E-03 0 % 

Evaporator insulation conductance UAins,e 5.474 W K-1 ±0.3052 W K-1 4.175 E-03 0 % 

42
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Table A-5: Uncertainty of cooling COP 

Total Partial Relative 

Parameter Symbol Value Uncertainty Derivative Error 

Cooling COP COP cool 2.844 ±0.01 

(0.35 %) -- --

Refrigerant press. at evaporator inlet P9 412 kPa ±3.5 kPa -1.232 E-02 39 % 

Refrigerant temp. at compressor discharge T3 81.97 °C ±0.09 °C 5.571 E-02 25 % 

Refrigerant temp. at compressor suction T1 12.73 °C ±0.09 °C 4.788 E-02 18 % 

Refrigerant press. at condenser exit P7 1104 kPa ±3.5 kPa 5.144 E-03 7 % 

Refrigerant temp. at expansion valve inlet T8 38.16 °C ±0.09 °C -2.746 E-02 6 % 

Refrigerant temp. at evaporator exit T11 12.09 °C ±0.09 °C 1.696 E-02 2 % 

Refrigerant evaporator diff. press. ∆Pe 43.92 kPa ±0.8 kPa 7.432 E-04 2 % 

Refrigerant condenser diff. press. ∆Pc 32.34 kPa ±1.3 kPa 4.731 E-04 1 % 

Refrigerant suction line diff. press. ∆Ps 14.22 kPa ±0.3 kPa 1.314 E-03 0 % 

Refrigerant mass flow mref ɺ 11.00 g s-1 ±0.016 g s-1 2.209 E-04 0 % 

A.11.2 Uncertainty Results Summary: Heating 

Table A-6 shows the dominant sources of instrument uncertainty for the refrigerant-side 

capacity are refrigerant mass flow (85 %), condenser exit temperature (9 %), and condenser inlet 

temperature (5 %). Table A-7 shows the dominant source of instrument uncertainty for the HTFside capacity is the HTF specific heat (97 %, estimate for instrument used to measure specific 

heat of water in the reference). Table A-8 shows the dominant sources of instrument uncertainty 

for cooling COP are evaporator inlet pressure (32 %), compressor discharge temperature (31 %), 

and compressor suction temperature (22 %). 

43
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Table A-6: Uncertainty of refrigerant-side condenser heating capacity
­

Total Partial Relative 

Parameter Symbol Value Uncertainty Derivative Error 

Refrigerant-side condenser capacity Qc ref , ɺ 1.873 kW ±0.004 kW 

(0.2%) -- --

Refrigerant mass flow mref ɺ 9.06 g s-1 ±0.016 g s-1 2.067 E-01 85 % 

Refrigerant temp. at condenser exit T7 37.39 °C ±0.09 °C -1.343 E-02 9 % 

Refrigerant temp. at condenser inlet T4 77.22 °C ±0.09 °C 9.443 E-03 5 % 

Refrigerant press. at condenser exit P7 992.8 kPa ±3.5 kPa -1.296 E-04 1 % 

Refrigerant condenser differential press. ∆Pc 27.2 kPa ±1.3 kPa -9.028 E-04 0 % 

Table A-7: Uncertainty of HTF-side condenser heating capacity 

Total Partial Relative 

Parameter Symbol Value Uncertainty Derivative Error 

±0.054 kW HTF-side condenser capacity Qɺ 1.852 kW -- -- c,HTF (2.9%) 

Condenser HTF capacity calibration cp HTF ,c 4.183 J g-1 K-1 ±3 % of cp 4.262 E-01 97 % uncertainty HTF ,c 

Condenser insulation conductance UAins,c 6.521 W K-1 ±0.4071 W K-1 1.056 E-02 1 % 

HTF mass flow mHTF ,c ɺ 98.81 g s-1 ±0.23 g s-1 1.804 E-02 1 % 

Condenser HTF thermopile voltage ∆VTP,c 2.318 mV ±0.007 mV 7.652 E-01 1 % 

Condenser HTF thermopile aTP,c,0 3.8494 E-02 ±5.3 E-04 4.057 E-03 0 % 

polynomial coefficients aTP,c,1 2.0241 E+03 ±1.5 E-01 1.025 E-03 0 % 

aTP,c,2 -5.5190 E+03 ±2.8 E+01 3.481 E-05 0 % 

aTP,c,3 3.8551 E+04 ±1.6 E+03 9.078 E-07 0 % 

aTP,c,4 -2.5440 E+05 ±2.9 E+04 2.309 E-08 0 % 

Temperature of ambient air 

surrounding condenser Tamb,c 24.3 °C ±0.6 °C -6.521 E-03 0 % 

Temperature at cold end of 

condenser HTF thermopile THTF,c,in 31.96 °C ±0.6 °C -4.093 E-03 0 % 

44
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Table A-8: Uncertainty of heating COP 

Total Partial Relative 

Parameter Symbol Value Uncertainty Derivative Error 

Heating COP COP heat 3.431 ±0.01 

(0.28 %) -- --

Refrigerant press. at evaporator inlet P9 314.8 kPa ±3.5 kPa -1.520 E-03 32 % 

Refrigerant temp. at compressor discharge T3 80.73 °C ±0.09 °C -5.933 E-02 31 % 

Refrigerant temp. at compressor suction T1 3.54 °C ±0.09 °C 5.000 E-02 22 % 

Refrigerant temp. at condenser inlet T7 37.39 °C ±0.09 °C -2.461 E-02 6 % 

Refrigerant press. at condenser exit P7 992.8 kPa ±3.5 kPa 5.532 E-04 5 % 

Refrigerant temp. at condenser inlet T4 77.22 °C ±0.09 °C 1.731 E-02 3 % 

Refrigerant condenser differential press. ∆Pc 27.2 kPa ±1.3 kPa 5.509 E-04 1 % 

Refrigerant suction line diff. press. ∆Ps 12.69 kPa ±0.3 kPa 1.520 E-03 0 % 

Refrigerant evaporator diff. press. ∆Pe 39.99 kPa ±0.8 kPa 1.520 E-03 0 % 

Refrigerant mass flow mref ɺ 9.06 g s-1 ±0.016 g s-1 2.187 E-04 0 % 

45
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Appendix B: Instrumentation with DAQ Number Convention
­

The instrument numbering scheme shown in Figure 2-1 and Figure 2-5 is based on the 

thermodynamic states 1 to 11. However, the raw data files recorded by the DAQ system use a 

different numbering scheme shown in Figure B-1 and Figure B-2. 

(a)	­ (b) 

Figure B-1:	­ Schematic of (a) condenser and (b) evaporator, showing the instrument 

numbering convention used in the DAQ system. 

46
­
Figure B-2: Schematic of MBBHP test apparatus including instrument numbering convention used in 

47 
 

 

 

 

 

 

 

 

 

 

 

Appendix C: Microfin Tube Surface Area
­

The microfinned surface area (per tube length) is estimated according to: 

Ai  sf 2e  = Nf + (C.1) L 

 cosα cosα cos(β 2)



  

where the variables represent: 

e = fin height 

Nf = number of fins 

sf = fin spacing 

α = fin helix angle 

β = fin angle 

and Ai/L is computed as 0.0424 m (1.67 in). The fin parameters are shown in Figure 2-5 and 

Table 2-2. 

48
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