IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 1 of 99

I. Executive Summary

Project Objective:

This project is an investigation of efficiency of R134a refrigerant systems for mobile air 

conditioning. The main objective for Team 2 is to evaluate the COP of R134a 

refrigerant systems with components and controls optimized for efficiency at moderate 

and low loads but with at least equal high load performance to the Phase II Alternate 

Refrigerant Cooperative Research Program [ARCRP] systems. The goal of Team 2 

was to demonstrate an improvement of 30% in system COP over the R134a System 

used in Phase II ARCRP. In the opinion of the OEM members of Team 2, this system 

represents efficiency significantly above the typical system used in production vehicles 

in 2005. Please note the discussion on pg 6 "Energy vs COP"; energy is a better metric 

because COP does not account for the change in magnitude of energy usage with

varying climatic conditions.

Project Overview: Purpose and Background:

The Improved Mobile Air Conditioning (IMAC) Cooperative Research Program (CRP)

began in August, 2004. The efficiency team's technical approach started with “brainstorming” ideas for potential technologies that would improve efficiency. This list was 

divided among the members and each member was asked to research individual 

technologies and assesses their potential benefits. This research was then discussed 

in the team meetings and a consensus was reached on the potential percentage 

improvement that each technology offered. Based on this assessment, the items were 

prioritized, and potential suppliers of these technologies were requested to submit 

sample components. This report presents the results of the evaluations that have been 

conducted on a System Test Stand in the lab and a demonstration vehicle.

The project began with organizational meetings in the Fall of 2004. The team then met 

regularly during 2005 and 2006 to review the progress of testing of the various 

technologies. An interim report was provided to the sponsors in December, 2005 and 

an update of this was presented in June, 2006. Vehicle development began in Spring,

2005 and continued through the end of 2006.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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The report is organized as follows:

I. Executive Summary

II. System Test Bench Evaluations

III. Vehicle Evaluations

Appendices:

IMAC2-A System setup tests

IMAC2-B B-o-B detail data

IMAC2-C Table of variables for bench testing

IMAC2-D Table of variables for the vehicle

IMAC2-E Vehicle data table

IMAC2-F Calculations for the bench test

IMAC2-G Vehicle COP and Capacity Calculations

IMAC2-H Details of algorithm for vehicle

IMAC2-I J2765 Procedure for measuring system COP [Coefficient of Performance] 

of a mobile air conditioning system on a test bench

IMAC2-J J2766 Standard for Annualized Climate Calculation of System Power 

Consumption of a mobile air conditioning system

IMAC2-K User’s manual for the vehicle controls

Summary of Results and Conclusions:

The reference baseline changed during testing and the technologies were not always 

additive in final Best of Best [B-o-B] system result. The team decided after analyzing 

the impact of COP as compared to energy consumption that it would be better to use 

compressor energy consumption as a measure of success. Comparing energy, the 

IMAC final best of best system run at 10°C evaporator out temperature demonstrated 

over 36% reduction in energy consumption as compared to the ARCRP baseline which 

was run at lower evaporator out temperatures. The COP improvement was 12.6% 

when compared at this same condition. At equal capacity, the COP improvement was 

15%. [The improvement is greater because the compressor efficiency is improved at 

higher displacement.] This is caused by the reduced efficiency of the variable 

compressor at reduced strokes. The raised evaporator temperature represents a

reduced reheat strategy.

ARCRP Base IMAC B-o-B

Teao= ARCRP

IMAC B-o-B

Teao= 10C

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

610,230 2.86 535,406 3.29 388,115 3.22

-12.3% 15.0% -36.4% 12.6%
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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At various intermediate steps, the efficiency improvement was measured against a 

baseline system that was assembled at the beginning of the IMAC project. This 

baseline was comprised of components from the previous Alternate Refrigerant 

Cooperative Research Program, but included a new condenser. The new condenser 

was configured by a new supplier to explore different subcooling pass configurations. 

The deliverables as measured against this original IMAC baseline system nearly met 

the program goals as shown in the table below. COP was much improved, and energy

consumption nearly met the 30% target. The original IMAC system was not as efficient 

as the ARCRP Phase II system. These results are from the technology screening test 

matrix that is less comprehensive than the full matrix used for the previous table.

IMAC Base 

Compr energy

IMAC B-o-B 

Compr energy

IMAC Base 

COP

IMAC B-o-B 

COP

663,495 481,791 2.20 3.39

27.4% 54.3%

Individual component technology changes showed even more improvement relative to 

the baseline IMAC system. 

Standards have been delivered that document the procedures for these measurements 

and annualization. A cost/benefit analysis has been completed to compare the 

technologies. 

Performance of this enhanced system has been demonstrated on a vehicle to match the 

current production system. An algorithm was developed and demonstrated on this 

vehicle that produced a real time calculation of system COP that could be used to 

optimize system controls for COP or capacity. The results of this algorithm correlated 

well with the bench test results.

Outside the original scope of the project, methods for measuring a fixed compressor 

with cycling clutch controls to a variable compressor have been developed. This 

demonstrated that variable compressors are more efficient than fixed compressors. 

Some new methods were developed to accurately measure the capacity of systems 

under cycling conditions. 

Below is a list of the chosen technologies and their relative values for both energy 

reduction, COP improvement, and a relative cost ranking as developed by IMAC Team 

2. Relative costs are the judgment of the Team members and are not based on any 

actual cost information. Energy improvement for some technologies can vary 

significantly depending on the baseline components (e.g. results for superheat control 

and internal heat exchangers can vary widely with different evaporators).
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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IMAC System Test Stand Results

Description of 

Technology

Annual COP 

Improvement 

[%]

Annual 

Energy 

Improvement 

[%]

Cost 

Ranking 

Low=1, 

High=10

COP/Cost Energy/Cost Comments

Improved 

Condensers

6.3 -7.4 4.0 1.6 -1.9

Improved 

Evaporators 21.2 -14.1 4.0 5.3 -3.5

Reduced Reheat 

[with External 

Controlled Variable 

Compressor]

-2.4 -24.1 6.7 -0.4 -3.6

Saves significant 

energy with a small 

decrease in COP –

shows that energy is 

the correct metric.

Internal Heat 

Exchanger 21.4 -12.3 6.5 3.3 -1.9

Results depend on 

evaporator design.

Cycling Fixed 

Compressors versus 

Variable Compressor 

Special study at 

request of Core Team 

– not considered for 

Best of Best

Fixed Piston

-18.7 +27.2 NA

Fixed Scroll

-11.1 +19.5 NA

Insulated Suction 

Line 1.0 -1.0 2.7 0.4 -0.4

Improvement was 

considered minimalonly two points run

Improved variable 

compressors, some 

with integrated oil 

separator

39.8 -18.3 6.7 5.9 -3.7 Limited test matrix

Condenser 

Subcooling Pass 

Design

6.3 -7.4 2.5 2.5 -3.0

Subcooling Control 

Algorithm 1.0 -1.0 7.5 0.1 -0.1

Improvement was 

considered minimalonly two points run

Superheat Control 

Algorithm [enabled by 

use of EXV]

28.6 -16.5 7.2 4.0 -2.3

Change from 7K to 1K 

at low loads. Cross 

charge TXV is similar. 

Results depend on 

evaporator design.

Best of Best System

versus ARCRPII 15.0 -36.4
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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Vehicle Results:

The main objective of the vehicle was to demonstrate that the improved control with 

optimized efficient components could be implemented into a production vehicle without 

degrading the maximum system performance. The graph below, taken from the final 

report of the 2006 Phoenix Alternate Refrigerant System Symposium, illustrates that the 

IMAC Optimized refrigerant vehicle did demonstrate equal capacity in the vehicle ride 

program. 

Panel Outlet Temperatures

ARSS 2006

49.7

54

57.7

55.8

56.8

55.7

58.5

37.3

43.7

46.3

43.3 43.8

45.3

46.8

35

40

45

50

55

60

65

Visteon Audi

R744

Hyundai - H Hyundai -

Prod

Honeywell 

Jetta

I-MAC / Load

reduction

I-MAC /

Optimized Ref. 

I-MAC /

Production. 

Degrees F.

City Traffic Road

Systems Technologies:

Except for the evaluation and comparison of the fixed compressor designs, all systems 

used a variable displacement compressor and a manually controlled electric actuated 

stepper motor expansion valves [EXV]. The systems evaluated here were composed of 

the following:

 Compressor

o IMAC Initial-Externally variable compressor

o Alternate Refrigerant CRP Phase II-same

o IMAC B-o-B-Improved Efficiency Externally variable compressor

 Evaporator

o IMAC Initial-Micro-channel design

o Alternate Refrigerant CRP Phase II-Enhanced Plate and Fin design

o IMAC B-o-B-Improved Efficiency -Micro-channel design

 Condenser-Micro-channel design, with varying subcooling tube configurations

o IMAC Initial-Micro-channel design

o Alternate Refrigerant CRP Phase II-Micro-channel design

o IMAC B-o-B-Improved Efficiency –Enhanced Micro-channel design
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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The reference baseline system for individual technologies was not always constant. 

The test matrix used to evaluate these individual technologies was a reduced matrix of 

the full IMAC test matrix and when the Annualization was conducted the results were 

not comparable to results of the full matrix. But, the relative rankings within each 

technology assessment should be the same. These technology assessments are 

discussed in more detail in section II.

Energy vs COP:

During the project development, it was recognized that energy is a better metric than 

COP based on life cycle analysis. Weighting COP gave too high a value to low load 

conditions where energy consumption is low. Energy consumption weighting better 

assesses the impact on the vehicle.

The steps to calculate include the following:

 Calculate average capacity of all tests of a particular technology

 Using COP, Calculate kW of energy consumption for the compressor for each 

point

 Using 300 hours of compressor usage/year, calculate kJ of energy consumption 

for the compressor

 Weight this using the speed and ambient profile agreed to by the team

Results presented above show both the COP and energy consumption annualized 

results. Ambient weighting is based on Visteon ambient analysis and assumes the AC 

is on above 10°C. The compressor RPM weighting is based on the weighting from a 

typical vehicle running on the combined FTP City and Highway driving cycles.

An example is shown below of the difference that was observed between two improved 

compressors [#4/#5] as compared to a baseline [#1], illustrating how COP and energy 

was weighted for COP and energy at each test point and then an average overall 

weighting.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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System Test Stand-Test Descriptions:

Tests included evaluation of cooling capacity and efficiency [COP] of each of the 

systems at four different speeds [900, 1500, 2500, 4000 RPM] and seven different load 

conditions. Comparisons were conducted at stabilized operation and equal capacity at 

the reduced load points as measured by the average evaporator out nozzle 

temperature. System capacity was targeted to maintain a 10C air outlet at these 

reduced load conditions. At the high load points, the compressor was set to maximum 

capacity. 

Test conditions were normalized in several aspects following the procedures 

established in the SAE ARCRP Phase II program prior to testing under stabilized 

conditions. Below is a summary of those areas evaluated:

1. Verify air velocity and temperature profiles at the inlet to each heat exchanger, 

and heat exchanger airside pressure drop.

2. Oil circulation was measured and adjusted per the compressor supplier 

recommendation prior to testing of the final configuration of components.

3. After the initial check, a system charge was selected per the charge 

determination procedure as outlined in the ARCRP Phase II final report. Details 

will be presented later in this report.

4. Charge determination

Standards:

The Efficiency Team developed a standard for the laboratory test method for testing 

mobile air conditioning systems for system efficiency. This team also developed an 

additional standard to standardize the method for Annualization of this energy 

consumption. Drafts of the two standards can be found in the appendix. They are as 

follows:

 J2765 Procedure for measuring system COP [Coefficient of Performance] of a 

mobile air conditioning system on a test bench

 J2766 Standard for Annualized Climate Calculation of System Power 

Consumption of a mobile air conditioning system

Cost vs Benefit summary:

The team has decided that the benefit part of the equation will be based on an 

annualized number for energy efficiency. The method used to determine this has been 

written into the SAE standard.

The cost portion of the equation was a consensus ranking of each of the technologies 

on a 1-10 scale as it is not possible to readily assess the specific implementation costs 

for each technology but the team feels we can assess the relative costs. The results 

are reported in the table above.

Acknowledgements:

Team two would first like to acknowledge the support of the sponsoring companies for 

funding of this Cooperative Research Program.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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We would like to acknowledge the support of the University of Illinois ACRC (Air 

Conditioning and Refrigeration Center) team members who completed the testing of the 

various system technologies during this program.

U of I Team Members

Pega Hrnjak, Professor 

Xinzhong Li, Research Scientist

Stefan Elbel, Graduate Research Assistant

Jing Che, Visiting Scholar

Radko Brock, Visiting Scholar

Jacob Virostko, Undergraduate Research Assistant

Team two would also like to acknowledge the donation of a Cadillac STS vehicle from 

General Motors for demonstration of the improved efficiency system and control 

algorithm.

Finally we would like to acknowledge all of the companies that contributed components 

for evaluation during the testing of these four alternative refrigerant systems. Due to 

confidentiality concerns, their names will not be disclosed here. There were some 

components that were submitted and not used based on preliminary evaluations.

Team Two Individual Members

Name Company

Jacob Bayyouk Sanden

David Barwin Nissan

Mahmoud Ghodbane Delphi

Hans Hammer Audi

John Meyer Visteon

Mark Spatz Honeywell

Steve Lepper Ford

Gene Dianetti Parker

Pankaj Mahajan DCX

Kwangtaek Hong DCX

Markus Wawzyniak Behr

Bill Hill GM

Dennis Littwin Fujikoki

Jan Zeng Denso
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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II. System Test Bench Evaluations

Technologies Considered:

Team 2 began by brainstorming a list of technologies for consideration. Technologies 

were assigned to team members to research. Team meetings, in 2004, reviewed the

potential benefits and ranked the technologies for COP improvement. This process 

resulted in the selection of the most promising technologies for testing. Below is the 

final evaluation matrix which lists the technologies considered and their assessment. 

Appendix A of the Dec, 2005 Interim Report includes the detailed assessments.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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1. Improved Heat Exchanger Designs (Condensers and Evaporators)

Presentations were given by Behr and Delphi on the opportunities for improvement. 

Based on these presentations, the team decided to pursue this technology for 

investigation. All suppliers were then requested to supply data on improved designs. 

This data was used by the OEM members to select heat exchanger designs for test. 

The specification allowed the option for thicker condensers and evaporators provided air 

side pressure drop was the same as baseline. 

2. External Control of Variable Compressors

The externally controlled variable compressor enables the system to raise the 

evaporator out pressure and correspondingly, the air discharge temperature to reduce 

the energy consumption of the compressor at low loads. The benefit for external control 

of evaporator outlet air temperature was not tested as a separate item. The individual

benefit was ranked based on the GM presentation, below is a summary:

 ARCRP data suggests a 10-20% COP improvement for external controlled 

compressor as compared to internal control.

 Supplier A estimated the savings to be 15-30% depending on loads.

 Supplier B estimated the savings to be 25-40% depending on loads.

The team agreed that 15% was a conservative estimate supported by ARCRP data. 

The Best of the Best I-MAC System was tested at elevated evaporator air out control 

temperatures and at the normal control temperatures used in ARCRPII.

3. Internal Heat Exchanger

This technology was presented by Visteon and supported by the University of Illinois. 

Initially, a commercial design was tested in two configurations to substantiate potential 

improvement. Subsequently, two automotive designs were tested. These tests

included studies of the impact of various superheat control strategies.

4. Scroll Compressor

Initially, the team decided not to pursue this technology based on discussion between a 

key supplier and the OEM’s. Later, in a Core Team meeting it was proposed that Team 

Two should take on the evaluation of fixed vs variable compressors. With this 

assignment, team two decided to pursue a comparison of fixed scroll vs fixed piston 

compressors as compared to a variable compressor baseline. 

5. Fixed vs Variable Compressor

A summary of the available public information was researched and presented by Ford

showing results from no difference to significant advantage for internal control variable. 

While this study was considered beyond the initial scope of I-MAC, it remains a subject 

of great interest given potential regulations and the lack of unanimous agreement. The 

Core team then requested that team two perform some evaluation of these 

technologies. Both fixed piston and scroll designs were tested. These tests required 

the implementation of new high speed data acquisition.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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6. Oil Separator

This technology is included in some of the optimized compressor efficiency designs and 

is included in item 12 below. It was not considered as a stand alone technology.

7. Ejector System

Ejector technology was presented by University of Illinois personnel. While there is 

potential benefit, the technology was rated to be a low priority due to higher complexity 

compared to other options.

8. Airflow Imbalance at condenser air inlet

This item was tabled as its benefits can be applied on all refrigerant systems. Airflow 

imbalance was evaluated in Phase I of the ARCRP and the impact on efficiency was not 

significant.

9. Increased use of Interior re-circulated air

This item was tabled as its benefits can be applied on all refrigerant systems. 

10. Reduced Front End Re-circulation

This item was tabled as its benefits can be applied on all refrigerant systems. The 

impact was studied with the final system configuration by running at elevated air inlet 

temperatures at the idle points to compare back to the performance of Phase II ARCRP.

11. Insulated Suction Line

This action was evaluated during the Best of the Best System test.

12. Optimized Compressor Efficiency

Compressor manufacturers were asked to provide prototype high efficiency externally 

controlled variable compressor designs. Two compressor suppliers provided both a 1st

generation and a 2nd generation of improved compressors for evaluation. Some 

designs include integrated oil separators.

13. Optimized System Control Strategy

GM presented ideas for controlling sub-cooling at lower evaporator loads (discussed in 

the results section in more detail). Varying sub-cooling with load was believed to offer 

meaningful COP improvement. The size of the condenser sub-cooling region was 

believed to be of importance. In preliminary testing, it was discovered that control of

evaporator out superheat also had a large impact on system efficiency at low loads. All 

of these items were included for testing and are discussed separately in the results 

section below.

14. Reduced Suction Line pressure drop

Parker Hannifin presented some lab test data on this item. The impact can be very 

dramatic. The baseline system was judged to already be on the low side of normal. 

Therefore, this item was not included for test.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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15. Flash gas removal

University of Illinois personnel indicated some potential benefit. While this technology is 

interesting, it seems to be more complicated and the team decided not to include in this 

evaluation.

16. Digital scroll

Nissan presented data on this subject. This technology applies only to electric driven 

compressors, and for this reason, this item was not included for test.

17. Improved evaporator air temperature distribution

All testing was done with uniform air temperature inlet to evaporator and condenser. 

The team consensus was this issue applies to all systems and there was not a separate 

evaluation of this effect during this research program. 

18. Matching components to compressor

The team decided that since matching system components is standard industry practice 

this item should not be part of the test program.

19. Electronic Expansion Valve

This technology is an enabler to provide control of evaporator out superheat. It was not 

tested as a separate item.

20. Electric Compressor

Since there are no (known) plans for high voltage on main stream vehicles, the team 

decided not to pursue testing.

21. Expander

This subject was researched by Honeywell and information was also received from 

Refrigeration Development Company Ltd. This technology was not deemed to provide 

sufficient improvement to warrant the added complexity.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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Technologies Tested:

Technologies selected for stand testing at the University of Illinois are shown in the 

following table.

Description of Technology

Improved Condensers

Improved Evaporators

External Controlled Compressor [Reduced Reheat]

Internal Heat Exchanger

Cycling Fixed Compressors versus Variable Compressor - NOT for 

BoB [Piston and Scroll type compressors]

Insulated Suction Line

Improved variable compressors, integrated oil separator

Condenser Subcooling Pass Design

Subcooling Control Algorithm

Superheat Control Algorithm w/EXV

Best of Best System

The improvement results for individual technologies are compared to a baseline system. 

Due to testing progression over a two year period, baselines for individual technologies 

are not the same. In some cases, the baseline changes within a technology test.

The Best of the Best (BoB) system includes:

 BoB Improved Condenser

 BoB Improved Evaporator

 External Controlled Compressor (elevated evaporator air off temperatures)

 BoB Improved Variable Compressor

 Superheat Control Algorithm w/ EXV

Baseline for the BoB system is the ARCRP Phase II R134a system.

The Fixed versus Variable compressor test was a special study at the request of the 

Core Team. Fixed compressors were not considered for the BoB system. As such, the 

fixed results will be presented separately after the other results.

Test Facility:

For continuity with the work conducted in ARCRP (Phases I and II), the ACRC (Airconditioning and Refrigeration Research Center) at the University of Illinois at UrbanaChampaign was selected for the SAE IMAC test activity. The following gives a brief 

summary of the test facility. 

The experimental facility, shown below, has two environmental chambers and a 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 14 of 99

compressor chamber in-between. Each of the two chambers contains a wind tunnel 

with variable speed blower and temperature controller, enabling a wide range of airflow 

rate and air temperature for the condenser and evaporator in the chambers. The 

evaporator chamber also has a steam supply and humidity controller to provide the 

latent load. Two chambers were designed to give three independent methods to 

determine system capacity: refrigerant side, airside, and calorimetric chamber balance. 

With the three methods used, instead of two as required by all applicable standards, the 

determination of system capacities is more reliable. In some cases, when the 

evaporator refrigerant exit is two-phase, or the system operates unstably (under some 

low load condition), only the airside balance is available. Careful design and calibration 

were made so that at least two of the three independent procedures could provide an 

agreement of ±3% as opposed to widely accepted standard of ±5%. 

The compressor chamber holds the compressor at a desired temperature to simulate 

temperature conditions in the vehicle engine compartment under the hood. Compressor 

power is obtained by shaft torque and speed measurement. A torque meter is located 

between the compressor and the driven clutch, eliminating belt and clutch losses. 

Evaporator

Chamber

Sc

Txri, Pxri

Tero

Trcpi, Prcpi

TC

W

H

B

Teai TG

Hu

N

Ten 1

Tdpen

Tdpei

 Peri

Dper

Sp

FS

Teao TG

Steam

M

Dpcn

Tcri

Dpcr

Ten 2

Dpen Dpea

Evap

CP

Mtr

 Tor

HG

Condenser

Chamber

Tgo

Tgi

B

H

CC

N

Tcn 1

TC

Tcai TG

Sp

W

Tcro,

Pcro

Tcao TG

M

Dpca

Tcn 2

Cond

Trcpo, Prcpo

Mg

TXV

Mr

Tcpc

TC

B - Blower, CP - Compressor, CC - Cooling Coil,

CH - Glycol Chiller,

Cond - Condenser, Dp - Differential Pressure Transducer, Evap - Evaporator, FS -

Flow Straightener, H - Heater, Hu - Humidifier, Mg - Glycol Mass Flow Meter, Mr - Refrigerant Mass Flow Meter, Mtr - Motor, N - Nozzle,

P - Pressure Transducer, RH - Relative Humidity Probe, Sc - Condensate Scale, Sp - Speed Controller and Tachometer,

T - Thermocouple,TC - Temperature Controller, TG - Thermocouple Grid, Tor - Torque Transducer, W - Watt Transducer, TXV

 Expansion Valve,

Indices : a - air, c - condenser, cp - compressor, e - evaporator, g - glycol, i - inlet, n - nozzle, o - outlet, r - refrigerant,

R/D - Receiver and Drier, Tdp - Dew Point Sensor, HG - Heat Gun, SG - Sight Glass

M - mixer,

CH

Compressor

Chamber

Sp

R/D

SG

 - Thermostatic

 x - expansion device

Schematic of the experimental facility
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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TM

Air in

Air in

SG

Mr

MF

IRD

Cond.

Evap.

Evaporator Chamber Wall Condenser Chamber Wall

Dpen

Comp.

Compressor Chamber

Motor

Tor

Tcao

Tcai

Dpcn Teai

Teao

Sp

Pcpro Pcpri

Tcpro Tcpri

Tcri Tcro

Pcro

TG

TG

TG

TG

DPcr

Tero

Pero

DPer

Ten 1

NZ

Ten 2

NZ

Tcn1 Tcn2

Tcph

Teri

EXV

Pxri

Txri

Instrumentation schematic
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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Methods of Capacity and COP Measurement:

Evaporator Capacity is measured using three independent methods:

 Air Capacity = dry air mass flow rate * air enthalpy difference – condensate 

enthalpy (small)

 Refrigerant Capacity = refrigerant mass flow rate * refrigerant enthalpy difference

 Chamber Capacity = electric heat – glycol cooling (required at low loads) + steam 

supplied energy – evaporator/cooling core condensate energy +/- heat gain/loss 

through the chamber walls (small)

Capacity Measurement Schematic

Reported Capacity, used to calculate COP, is the average of Air Side and Chamber 

Capacity EXCEPT where noted in the results. At low loads, Air Side Capacity is used 

because it is the only reliable method.

Compressor Power is calculated from measurement of shaft torque and speed. A strain 

gage torque meter is mounted in line with the compressor through a flexible coupling. 

Speed (rpm) is measured with an optical pulse counter. 

Coefficient of Performance = Evaporator Capacity (kW) / Compressor Power (kW).

Power for the condenser/gas cooler fan and evaporator blower was not included in the 

COP calculations. Heat exchanger dry air pressure drops were close to each other. 

Therefore, power for the fan and blower would be similar between the systems.

Testing fixed compressors during cycling required high speed data (10 Hz) acquisition. 

Power measurement was a straight forward high speed data acquisition and integration 

of torque and rpm. Capacity measurement, however, presented an issue. During 

cycling, refrigerant enthalpy is not accurate. As noted above, at the lowest loads only 

Air Side Capacity is reliable. For cycling conditions, dry bulb and humidity leaving the 

evaporator were integrated separately from the high speed data to calculate sensible 

and latent capacities. The issue encountered was that the humidity measurement used

air mass flow rate

hai=h (T, RH) hao=h (T, RH)

refrigerant mass flow rate

hro=h (T, P) hri=h (T, P, x)

±

steam

glycol

air mass flow rate

hai=h (T, RH) hao=h (T, RH)

refrigerant mass flow rate

hro=h (T, P) hri=h (T, P, x)

±

steam

glycol

condensate
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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a chilled mirror which is a slow response device itself and the air sample must travel 

some distance into the device. The chilled mirror is highly accurate at the high relative 

humidity leaving the evaporator. A review of faster response devices revealed that 

none are sufficiently accurate at these high humidity conditions. For the conditions 

where Chamber Capacity is available, a comparison was made with Air Side Capacity 

measure as just described. After the system stabilized to repetitive cycles, data was 

taken and integrated over at least 10 whole cycles. Accuracy assessment is below.

Overall, the assessment of results indicated the following:

 Capacity accurate to ±3% at loads greater than 1500W.

 Capacity accurate to ±100W at loads less than 1500W.

 Compressor power accurate to ±20W.

 COP accurate to ±3% at loads greater than 1500W.

 COP accurate to ±9% at loads less than 1500W.

Selected error analysis of Capacity and COP accuracy are shown in the graphs below 

for both fixed and variable compressor. 

SAE IMAC fixed compressor cooling capacity calculation: Error = (Qe_air - Qe_ch) / Qe_ch

Note: Qe_air is calculated based on fast sampling (dynamic) data 

Qe_ch is calculated based on slow sampling (stable) data 

-4.95

-3.81

3.18

5.59 5.07

20.59

-10

-5

0

5

10

15

20

25

H45 cycling H35a cycling I25a cycling L25a cycling M25a cycling H25a cycling

Test condition

Error [%]

Error: Qe_air compared with Qe_ch

Frosting occurred so chamber balance was off
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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SAE IMAC variable compressor with TXV cooling capacity calculation: 

Error = (Qe_air - Qe_ch) / Qe_ch

Note: Both Qe_air and Qe_ch are calculated based on slow sampling (stable) data 

-2.60 -2.48

1.06

3.83

4.47 4.34

-3

-2

-1

0

1

2

3

4

5

H45 cycling H35a cycling I25a cycling L25a cycling M25a cycling H25a cycling

Test condition

Error [%]

Error: Qe_air compared with Qe_ch

Details of Test Procedures for Capacity and Energy:

All tests for system COP and component performance were done at steady state 

(stable) conditions. The SAE ARCRP Phase II test matrix was reduced to more 

efficiently screen various technologies under consideration. The intention is to run 

complete SAE ARCRP Phase II test matrix with the combination of technologies that 

provides the optimal system efficiency. The test matrix is shown in Table III-2 below. 

These steady state conditions cover a significant range of typical vehicle air conditioning 

operation. 

Systems were compared at equal capacity by controlling evaporator air outlet 

temperature as measured at the airflow nozzles at reduced load points. System 

capacity was controlled by varying compressor displacement. The evaporator air outlet 

temperature was controlled to 10oC. The compressor was run at full capacity at 35°C 

and 45°C test ambient conditions. 

Compressor chamber temperature set to 30K above condenser air inlet temperature.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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Controls Discussion:

For variable compressor operation:

The compressor was controlled with a PWM [Pulse Width Modulated] control. This was 

used to allow control of the displacement so points with equal evaporator out air 

temperature could be run. The displacement control was based on air out of evaporator 

[actually the average of the two thermocouples located in the nozzle downstream of the 

evaporator]. The expansion valve controlled superheat in the range of 6-8K at high load 

conditions and 2-4K at low load conditions. 

For fixed compressor operation:

For the fixed compressor system, besides the change of compressor, the EXV was

replaced by a TXV as it is difficult to control Evaporator superheat during cycling with a 

manual EXV. The maximum cycle rate was specified as 6 cycle per minute for the fixed 

compressors running at cycling mode. The evaporator out air temperature (measure at 

nozzle) was used as the input signal for a PID controller to control compressor clutch. 

The temperature variation band was specified as ±1C. However, this band was allowed 

to be widened up to ±3C (or other number) to keep the cycle rate lower than 6 

cycles/minute.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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Table III-2 SAE IMAC test matrix (reduce matrix only for component performance 

comparison test)

Test Name

Ambient 

temp. 

[

oC/o

F]

Compr 

speed 

[rpm]

Condenser Evaporator

Inlet air 

temp. 

[

oC/ o

F]

Flow rate 

[l/s/CFM] Inlet air 

temp. 

[

oC/ o

F]

Flow 

rate 

[l/s/CFM

]

Simulated 

air 

selection

RH 

[%]

Dew 

point 

[

oC]

I60a 45/113 900 60/140 236/500 35/95 109/230 Recirc 25 12

I60b “ 900 60/140 342/725 “ 109/230 “ “ “

I60c “ 900 60/140 472/1000 “ 109/230 “ “ “

I45 “ 900 45/113 342/725 “ 109/230 “ “ “

L45 “ 1500 “ 689/1460 “ 130/275 “ “ “

M45 “ 2500 “ 1067/2260 “ 130/275 “ “ “

H45 “ 4000 “ 1067/2260 “ 130/275 “ “ “

I50a 35/95 900 50/122 342/725 35/95 109/230 OSA 40 19.4

I35a “ 900 35/95 342/725 “ 109/230 “ “ “

L35a “ 1500 “ 689/1460 “ 130/275 “ “ “

M35a “ 2500 “ 1067/2260 “ 130/275 “ “ “

H35a “ 4000 “ 1067/2260 “ 130/275 “ “ “

I35b 35/95 900 35/95 342/725 35/95 90/190 OSA 40 19.4

L35b “ 1500 “ 689/1460 “ 60/127 “ “ “

M35b “ 2500 “ 1067/2260 “ 60/127 “ “ “

H35b “ 4000 “ 1067/2260 “ 60/127 “ “ “

I40a 25/77 900 40/104 342/725 25/77 90/190 OSA 80 21.3

I25a “ 900 25/77 342/725 “ 90/190 “ “ “

L25a “ 1500 “ 689/1460 “ 60/127 “ “ “

M25a “ 2500 “ 1067/2260 “ 60/127 “ “ “

H25a “ 4000 “ 1067/2260 “ 60/127 “ “ “

I40b 25/77 900 40/104 342/725 25/77 90/190 OSA 50 13.9

I25b “ 900 25/77 342/725 “ 90/190 “ “ “

L25b “ 1500 “ 689/1460 “ 60/127 “ “ “

M25b “ 2500 “ 1067/2260 “ 60/127 “ “ “

H25b “ 4000 “ 1067/2260 “ 60/127 “ “ “

I25c 25/77 900 25/77 342/725 25/77 60/127 OSA 50 13.9

L25c “ 1500 “ 689/1460 “ 40/85 “ “ “

M25c “ 2500 “ 1067/2260 “ 40/85 “ “ “

H25c “ 4000 “ 1067/2260 “ 40/85 “ “ “

I30 15/59 900 30/86 342/725 15/59 60/127 OSA 80 11.6

I15 “ 900 15/59 342/725 “ 60/127 “ “ “

L15 “ 1500 “ 689/1460 “ 40/85 “ “ “

M15 “ 2500 “ 1067/2260 “ 40/85 “ “ “

H15 “ 4000 “ 1067/2260 “ 40/85 “ “ “

Yellow background for Reduced Matrix
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 21 of 99

The reduced test matrix for fixed compressors included selected 4000 rpm points. Each 

fixed compressor had its speed adjusted to match capacity, on average, at I35a, L35a, 

and M35a.

Baseline Variable Piston [#4] Fixed Piston [#8] [1.0] Fixed Scroll [#7] [1.65]

900 900 1485

1500 1500 2475

2500 2500 4125

4000 4000 6600

Annualization procedure:

For each individual technology, the annualization weighting factors were adjusted based 

on the points that were run for each evaluation. In running the final Best-of-Best 

evaluation with a full test matrix, it was discovered that the absolute annual savings was 

over estimated as a result of doing this evaluation utilizing the reduced test matrix. 

However, the relative ranking of the technologies should still be valid.

Table 5 Ambient weighting profile based on Visteon Climate Model

Outdoor load condition

Temp/ humidity

Weighting for USclimate

<15°C 10.6%

15°C/ 80% 46.2%

25°C/ 50% low 14.2%

25°C/ 50% high 14.2%

25°C/ 80% high 9.5%

35°C/ 40%;low 1.3%

35°C/ 40%;high 3.8%

45°C/ 25% 0.2%

Ignore I60 "a" and "c" 89.4%

Weight elevated idle 33% [only 

I60b]

Non elevated 67%

Table 6 Compressor Speed profile

Compressor RPM Percent Time at RPM

USA

<1000 17%

1500 43%

2500 35%

>4000 5%
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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System Setup Test Procedures:

Inlet air velocity and temperature distribution:

These tests were performed prior to starting the refrigeration system performance test 

to verify uniformity of air velocity and temperature at the inlet to the heat exchangers. 

Each test was performed for both heat exchangers (condenser and evaporator) and 

over a relevant range of air velocities and inlet air temperatures. 

For each of the system tested, the adjustment of air velocity distribution was not 

straightforward. After several iterations, condenser/gas cooler velocity profiles were 

achieved within ±8% for each of the predefined three airflow conditions and the 

evaporator velocity profiles were measured within ±6% for each of predefined two 

airflow conditions. The results at selected airflow conditions are presented below. 

Examples of Inlet air velocity and temperature profiles are shown below:

SAE IMAC R134a evaporator inlet temperature profile

20.0

25.0

30.0

35.0

40.0

45.0

50.0

T11 T12 T13 T14 T15 T21 T22 T23 T24 T25 T31 T32 T33 T34 T35 T41 T42 T43 T44 T45 T51 T52 T53 T54 T55

TC number

Temperature [oC]

Evaporator inlet air temperatures
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 23 of 99

34 100 166 232

31

94

157

220

cm

cm

Evaporator Velocity Distribution at 190cfm

0.78-0.80

0.76-0.78

0.74-0.76

0.72-0.74

Evaporator inlet air velocities

SAE IMAC R134a condenser inlet temperature profile

20.0

25.0

30.0

35.0

40.0

45.0

50.0

T11 T12 T13 T14 T15 T21 T22 T23 T24 T25 T31 T32 T33 T34 T35 T41 T42 T43 T44 T45 T51 T52 T53 T54 T55

TC number

Temperature [oC]

Condenser inlet air temperatures
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 24 of 99

5 159 313 467 621

33

113

193

273

353

mm

mm

Evaporator Velocity Distribution at 1460cfm

2.30-2.40

2.20-2.30

2.10-2.20

2.00-2.10

Condenser inlet air velocities
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 25 of 99

Air side pressure drop:

Each time there was a heat exchanger change, the dry air pressure drop was measured 

at room conditions to verify the specification. The airside pressure drop of both the 

evaporator and the condensers are shown below. 

SAE IMAC R134a evaporator airside pressure drop

0

20

40

60

80

100

120

140

160

180

0.00 0.50 1.00 1.50 2.00 2.50 3.00 3.50 4.00

Face Velocity [m/s]

Airside pressure drop DPea [Pa]

Phase II R152a/R134a Evap

IMAC Evap#1

IMAC Evap#2

IMAC Evap#3

Evaporator air side pressure drop comparison

SAE IMAC R134a condenser airside pressure drop vs. face velocity

0

20

40

60

80

100

120

140

160

180

200

0 1 2 3 4 5 6 7

Face velocity [m/s]

Airside pressure drop DPca [Pa]

Phase II R152a/R134a

IMAC R134a cond#3

IMAC R134a cond#6

IMAC R134a cond#7

Comparison of air side pressure drop of condensers

[Note: Condensers #1/5 have the same air side design as #3]
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 26 of 99

Refrigerant charge determination:

Charge determination was conducted at three different conditions (I35, L35 and I25) 

from the test matrix presented later in this chapter. Evaluation at three conditions 

allowed evaluation of optimum charge at low and high ambient conditions. In some 

cases only one condition was run to check charge where it was expected that there 

would be no impact on charge. Optimum charge was picked based on standard 

industry practices.

SAE IMAC charge determination test matrix

Test 

Name

Ambient 

temp. 

[

oC/o

F]

Compr 

speed 

[rpm]

Condenser Evaporator

Inlet air 

temp. 

[

oC/ o

F]

Flow rate 

[l/s/CFM]

Inlet air 

temp. 

[

oC/ o

F]

Flow rate 

[l/s/CFM]

Simulated 

air 

selection

RH 

[%]

Dew 

point 

[

oC]

I35a 35/95 900 35/95 342/725 35/95 109/230 OSA 40 19.4

L35a 35/95 1500 35/95 689/1460 35/95 130/275 OSA 40 19.4

I25a 25/77 900 25/77 342/725 25/77 90/190 OSA 80 21.3

The charging procedure was as follows: after evacuation, the system was charged with 

the lowest reasonably expected quantity of refrigerant 500~600g. The system was 

maintained at constant air temperature, humidity, and airflow rates through condenser 

and evaporator, and compressor speed, as prescribed by the charging condition. When 

steady state conditions were achieved data were taken and refrigerant was added in 

increments of 50g (or 25g). An example figure below shows charge determination test 

results for one of the systems tested.

SAE IMAC IHX1 I35a charge test

0

2

4

6

8

10

12

14

16

18

900 1000 1100 1200 1300 1400 1500

Charge [g]

Temp [C] / DT [C] / Q [kW] / COP [-]

0

200

400

600

800

1000

1200

1400

1600

1800

P [kPa]

COP_indoor_avg

DTsub_cro

Qindoor_avg

Ten

Selected charge

= 1300 [g]

Prcpi

Prcpo
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 27 of 99

Oil circulation rate measurement:

Before the final performance test, oil circulation rate was measured with two different 

compressor speeds and conditions. The oil type, the initial oil charge, and the target oil 

circulation values were recommended by the compressor manufacturer. Oil circulation 

was measured by taking refrigerant/oil samples from the liquid line and measuring the 

mass ratio of refrigerant to oil. 

The measurement results for oil circulation rate and system charge amounts for each 

system are given in the following table. 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 28 of 99

Description of Technology Charge 

Amount [gm] Oil Circulation (%) Comments

900 rpm 4000rpm

Improved Condensers

Baseline (Cond #3) 1050 1.46 5.95 Oil charge 137 cc, comp#4

Condenser #6 1200 - -

No oil circulation rate were measured for 

cond #6 and #7 but with same oil charge 

137 cc as the baseline

Condenser #7 1200 - - Oil charge 137 cc

Improved Evaporators

Baseline (Evap #1) 1050 1.46 5.95 Oil charge 137 cc

Evaporator #2 950 - - Oil charge 137 cc

Evaporator #3 1050 - - Oil charge 137 cc. 

Evaporator #4 1025 - - Oil charge 137 cc

Internal Heat Exchangers

Baseline 1100 0.33% 0.87% comp#2

Commercial 1 HX #1 1150 0.82% -

Commercial 2 HX in Parallel #2 1250 1.05% 

(1500rpm) -

Baseline 1200 1.46 5.95 Oil charge 137 cc, comp#4

Automotive Tube Type #4 1300 1.84 3.85 

(2500rpm) Oil charge 137 cc, comp#4

Automotive Core Type #4 1325 - - Oil charge 137 cc, comp#4

Cycling Fixed Compressors:

Baseline (BoB Variable) 1150 1.46 5.95 Oil charge 137 cc, with TXV

Fixed Piston comp#8 1150 0.93 1.25 

(2500rpm)

Fixed Scroll comp#7 1150 8.2

(1485 rpm) -

Insulated Suction Lines

Tested in BoB System with IHX 

Type #4 1325 - - Oil charge 137 cc, comp#4

Improved variable compressors

Baseline comp#1 1350 0.73 1.22

Compressor #2 1350 0.23 0.87 Integrated oil separator

Compressor #3 1350 0.41 1.42

Compressor #4 1150 1.46 5.95

Compressor #6 1150 1.52 2.92

Condenser Subcooling Pass 

Designs:

Design #1 1350 0.73 1.22 External receiver

Design #3 1250 ‘- ‘- Integrated receiver

Design #5 1300 ‘- ‘- Integrated receiver

Best of Best System

ARCRPII Baseline 1125 0.66 1.16

IMAC BoB 1150 1.46 5.95
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 29 of 99

Capacity and Energy Test Results:

Improved Condensers:

Comparison of IMAC heat exchangers to industry average:

Condenser Face Area

0.1

0.2

0.3

0.4

Face Area m

2

Europe Japan U.S. Group Average

0,247 m2 avg.all systems

ARCRP Phase I 

Systems

ARCRP Phase II 

Systems

IMAC system

241,600

The OEM team members developed a condenser component specification and invited 

condenser manufacturers to submit designs. The specification allowed for a thicker 

core than baseline so long as air side pressure drop specification was not exceeded.

Condenser designs were submitted by four suppliers. The OEM members selected the 

two highest performance designs including one that is thicker than the baseline.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 30 of 99

SAE IMAC system component table for condenser study

Served as baseline

Component Feature Cond#3 Cond#6 Cond#7

Condenser

Core effective face [mm] W626xH386XD16 W580xH418xD16 W546xH396xD22

Face area [mm2]

 241,636 242,440 216,216

Core volume [mm3] 3,866,176 3,879,040 4,756,752

Aspect ratio [-] 0.62 0.72 0.73

No. of tubes 39 (micro-channel) 42 (microchannel)

61 (microchannel)

No. of Passes 3 (22-11-6) 3 (?-?-?) 4 (29-18-8-6)

Mass [kg] 2.1 2.5 2.5

Evaporator (Evap#3)

Core effective face [mm] W267xH257xD50

Face area [mm2]

 68,619

Core volume [mm3] 3,430,950

Aspect ratio [-] 0.96

Type (No. of tubes) micro-channel

No. of Passes

Mass [kg] 1.6 (est.)

Compressor (Comp#4)

Type External controlled variable displacement

Displacement [cc] 162

Mass (including pulley) [kg] 6.5 (est.)

Receiver/Drier Mass [kg] IRD, see condenser

Total Mass [kg] 10.2 10.6 10.6

Note:

1. Three condensers were tested and all have an IRD. Cond#3 was the baseline.

2. Cond#7 was chosen for the final system.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 31 of 99

Following are the results of the condenser assessments:

SAE IMAC R134a condenser airside pressure drop vs. face velocity

0

20

40

60

80

100

120

140

160

180

200

0 1 2 3 4 5 6 7

Face velocity [m/s]

Airside pressure drop DPca [Pa]

Phase II R152a/R134a

IMAC R134a cond#3

IMAC R134a cond#6

IMAC R134a cond#7

SAE IMAC Condenser comparison: Evaparator Capacity Qe

Note: (1) Maximum Qe for 35degC points and 

(2) controlled Ten=10degC for 25degC points 

0

1

2

3

4

5

6

7

8

I45 L45 I35a M35a I25a M25a

Test condition

Qe [kW]

Cond#3 baseline

Cond#6

Cond#7
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 32 of 99

SAE IMAC Condenser comparison: COP

Note: (1) Maximum Qe for 35degC points and 

(2) controlled Ten=10degC for 25degC points 

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

I45 L45 I35a M35a I25a M25a

Test condition

COP [-]

Cond#3 baseline

Cond#6

Cond#7

SAE IMAC Condenser comparison: Wcomp

Note: (1) Maximum Qe for 35degC points and 

(2) controlled Ten=10degC for 25degC points 

0

0.5

1

1.5

2

2.5

3

3.5

4

I45 L45 I35a M35a I25a M25a

Test condition

Wcomp [kW]

Cond#3 baseline

Cond#6

Cond#7
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 33 of 99

SAE IMAC Condenser comparison: Pcro

Note: (1) Maximum Qe for 35degC points and 

(2) controlled Ten=10degC for 25degC points 

0

200

400

600

800

1000

1200

1400

1600

1800

2000

I45 L45 I35a M35a I25a M25a

Test condition

Pcro [kPa]

Cond#3 baseline

Cond#6

Cond#7

Only the following points were run:

I45, L45, I35a, M35a, I25a, M25a

Weighting was adjusted to weight only these points. Weighting of 15 C was 

assigned to 25 C.

Cond #3 Cond #3 Cond #6 Cond #6 Cond #7 Cond #7

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

kJ kJ kJ

918,844 3.77 914,764 3.74 850,908 4.00

Energy 

Comparison

-0.4% -0.8% -7.4% 6.3%

COP 

comparison:
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 34 of 99

Improved Evaporators:

Evaporator Face Area

0.04

0.05

0.06

0.07

0.08

Face Area m

2

Europe Japan U.S. Group Average

0,0572 m2 avg. all systems

ARCRP Phase II 

Systems

IMAC system

66,766

ARCRP Phase I 

Systems

The OEM team members developed an evaporator component specification and invited 

evaporator manufacturers to submit designs. The specification allowed for a thicker 

core than baseline so long as air side pressure drop specification was not exceeded. 

Two suppliers submitted evaporator designs. Since there were only two submissions, 

the OEM members decided to test both including one that is thicker than the baseline.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 35 of 99
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 36 of 99

SAE IMAC system component table for evaporator study

Served as baseline

Component Feature IMAC Evap#1 IMAC Evap#2 IMAC Evap#3

Condenser (Cond#3)

Core effective face [mm] W626xH386XD16

Face area [mm2]

 241,636

Core volume [mm3] 3,866,176

Aspect ratio [-] 0.62

No. of tubes 39 (micro-channel)

No. of Passes 3 (22-11-6)

Mass [kg] 2.1

Evaporator

Core effective face [mm] W266xH251xD38 W270xH235xD35 W267xH257xD50

Face area [mm2]

 66,766 63,450 68,619

Core volume [mm3] 2,537,108 2,220,750 3,430,950

Aspect ratio [-] 0.94 0.87 0.96

Type (No. of tubes) micro-channel (39) micro-channel micro-channel

No. of Passes two slabs, passes unknown two passes two slabs, passes unknown

Mass [kg] 1.2 1.1 1.6 (est.)

Compressor (Comp#4)

Type External controlled variable displacement

Displacement [cc] 162

Mass (including pulley) [kg] 6.4

Receiver/Drier Mass [kg] IRD, see condenser

Total Mass [kg] 9.7 9.6 10.1

Note:

1. Four evaporators were tested. Evap#1 was the baseline.

2. Evap#3 was chosen for the final system.

SAE IMAC R134a evaporator airside pressure drop

0

20

40

60

80

100

120

140

160

180

0.00 0.50 1.00 1.50 2.00 2.50 3.00 3.50 4.00

Face Velocity [m/s]

Airside pressure drop DPea [Pa]

Phase II R152a/R134a Evap

IMAC Evap#1

IMAC Evap#2

IMAC Evap#3
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 37 of 99

SAE IMAC Evaporator comparison: Evaparator Capacity Qe

Note: (1) Maximum Qe for 35degC points and 

(2) controlled Ten=10degC for 25degC points 

0

1

2

3

4

5

6

7

8

I35a M35a I25a M25a

Test condition

Qe [kW]

Evap#1 baseline

Evap#2

Evap#3

SAE IMAC Evaporator comparison: COP

Note: (1) Maximum Qe for 35degC points and 

(2) controlled Ten=10degC for 25degC points 

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

I35a M35a I25a M25a

Test condition

COP [-]

Evap#1 baseline

Evap#2

Evap#3
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 38 of 99

SAE IMAC Evaporator comparison: Wcomp

Note: (1) Maximum Qe for 35degC points and 

(2) controlled Ten=10degC for 25degC points 

0

0.5

1

1.5

2

2.5

3

3.5

4

I35a M35a I25a M25a

Test condition

Wcomp [-]

Evap#1 baseline

Evap#2

Evap#3

Evap#1 temp distribution

0

2

4

6

8

10

12

14

16

18

20

Teao1

Teao11Teao12Teao13Teao14

Teao2

Teao21Teao22Teao23Teao24

Teao3

Teao31Teao32Teao33Teao34

Teao4

Teao41Teao42Teao43Teao44

TC number

Temp [degC]

I35a

M35a

I25a

M25a
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 39 of 99

Evap#2 temp distribution

0

2

4

6

8

10

12

14

16

18

20

Teao1

Teao11Teao12Teao13Teao14

Teao2

Teao21Teao22Teao23Teao24

Teao3

Teao31Teao32Teao33Teao34

Teao4

Teao41Teao42Teao43Teao44

TC number

Temp [degC]

I35a

M35a

I25a

M25a

Evap#3 temp distribution

0

2

4

6

8

10

12

14

16

18

20

Teao1

Teao11Teao12Teao13Teao14

Teao2

Teao21Teao22Teao23Teao24

Teao3

Teao31Teao32Teao33Teao34

Teao4

Teao41Teao42Teao43Teao44

TC number

Temp [degC]

I35a

M35a

I25a

M25a
IMAC Team 2 - Improved Efficiency Final Report April, 2007

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Only the following points were run:

I35a, M35a, I25a, M25a

Weighting was adjusted to weight only these points. Weighting of 15 C was 

assigned to 25 C. Weighting of 45 C was assigned to 35 C.

Evap#1 Evap#1 Evap#2 Evap#2 Evap#3 Evap#3

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

kJ kJ kJ

1,046,341 3.11 964,406 3.50 899,173 3.77

Energy 

Comparison

-7.8% 12.5% -14.1% 21.2%

COP 

comparison:
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 41 of 99

Internal Heat Exchangers w/ superheat optimization:

In ARCRP, traditional co-tube designs were tested with R134a and they showed little 

improvement. The limitation is limited heat exchanger effectiveness combined with 

added suction line pressure drop. Also in ARCRP, a commercial design was tested with 

an alternative refrigerant and it showed more promise. For I-MAC, the team first tested 

a single commercial IHX and two commercial IHXs in parallel for proof of concept.

IHX Study one:

SAE IMAC system component table for IHX study: system No.1

Served as baseline

Component Feature without IHX IHX#1 IHX#2

Condenser (Cond#3)

Core effective face [mm] W626xH386XD16

Face area [mm2]

 241,636

Core volume [mm3] 3,866,176

Aspect ratio [-] 0.62

No. of tubes 39 (micro-channel)

No. of Passes 3 (22-11-6)

Mass [kg] 2.1

Evaporator (Evap#1)

Core effective face [mm] W266xH251xD38

Face area [mm2]

 66,766

Core volume [mm3] 2,537,108

Aspect ratio [-] 0.94

No. of tubes 39 (micro-channel)

No. of Passes two slabs, passes unknown

Mass [kg] 1.2

Compressor (Comp#2)

Type External controlled variable displacement

Displacement [cc] 160

Mass (including pulley) [kg] 6.1

Receiver/Drier Mass [kg] IRD, see condenser

IHX Mass [kg] N/A 1.8 2.8

Total Mass [kg] 9.4 11.2 12.2

Note:

1. Two IHXs were tested with this system.

2. IHX#1 was a single Danfoss IHX (1.83kg). 

3. IHX#2 was two Danfoss IHX connected in parallel (2.8 kg).
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 42 of 99

SAE IMAC IHX Study: Suction line DP comparison

20.0

27.1

9.9

7.9

2.8

0.4

37.8

57.7

18.5

9.1

22.1

33.6

12.1

5.0

1.3 0.6

0

10

20

30

40

50

60

70

L35a M35a I25a M25a I15 M15

Test conditon

DP [kPa]

Baseline data

1IHX

2IHX

SAE IMAC IHX Study: Cooling capacity Qe

Note: cooling capacity was controlled to be equal as the baseline. Only M25a is an exception (Baseline Ten=8.8C, IHX Ten=10C).

6.20

6.92

4.25

3.00

0.64

0.43

6.11

6.66

4.31

2.65

6.11

6.84

4.32

2.62

0.57 0.40

0

1

2

3

4

5

6

7

8

L35a M35a I25a M25a I15 M15

Test conditon

Qe [kW]

Baseline data

1IHX

2IHX
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 43 of 99

SAE IMAC IHX Study: COP comparison

2.71

2.08

3.42

2.89

3.17

1.49

2.71

1.98

3.67

3.48

2.77

2.17

3.62 3.60

5.61

1.66

0

1

2

3

4

5

6

L35a M35a I25a M25a I15 M15

Test conditon

COP [-]

Baseline data

1IHX

2IHX

SAE IMAC IHX Study: IHX effectiveness comparison

0.32 0.31 0.29

0.36 0.37 0.37

0.42

0.47

0.67

0.73

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

L35a M35a I25a M25a I15 M15

Test conditon

Eta_ihx [-]

1IHX 2IHX
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 44 of 99

SAE IMAC IHX Study: Wcomp comparison

2.29

3.32

1.25

1.04

0.20

0.29

2.25

3.36

1.18

0.76

2.15

3.09

1.19

0.75

0.10

0.24

0

0.5

1

1.5

2

2.5

3

3.5

4

L35a M35a I25a M25a I15 M15

Test conditon

Wcomp [kW]

Baseline data

1IHX

2IHX

Only the following points were run:

L35a, M35a, I25a, M25a, I15, M15

Weighting was adjusted to weight only these points. Weighting of 45 C was 

assigned to 35 C.

Baseline Baseline 1 IHX 1 IHX 2 IHX 2 IHX

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

kJ kJ

778,540 2.28 697,199 2.75 683,093 2.77 Energy Comparison

-10.4% 20.6% -12.3% 21.4% COP comparison:
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 45 of 99

IHX Study two:

Test results for the two commercial IHXs in parallel were good, so the team decided to 

test two designs specifically developed for automotive application. One is an enhanced 

co-tube design. The second is a compact heat exchanger design.

SAE IMAC system component table for IHX study: system No.2

Served as baseline

Component Feature Baseline - without IHX IHX#2 IHX#3 IHX#4

Condenser (Cond#7)

Core effective face [mm] W546xH396xD22

Face area [mm2]

 216,216

Core volume [mm3] 4,756,752

Aspect ratio [-] 0.73

No. of tubes 61 (micro-channel)

No. of Passes 4 (29-18-8-6)

Mass [kg] 2.5

Evaporator (Evap#3)

Core effective face [mm] W267xH257xD50

Face area [mm2]

 68,619

Core volume [mm3] 3,430,950

Aspect ratio [-] 0.96

No. of tubes micro-channel

No. of Passes two slabs, passes unknown

Mass [kg] 1.6 (est.)

Compressor (Comp#4)

Type External controlled variable displacement

Displacement [cc] 162

Mass (including pulley) [kg] 6.5 (est.)

Receiver/Drier Mass [kg] IRD, see condenser

IHX Mass [kg] N/A 2.8 0.4 0.8

Total Mass [kg] 10.5 13.4 11 11.4

Note:

1. Three IHXs were tested with this system.

2. IHX#2 was two Danfoss IHX connected in parallel (2.8 kg).

3. IHX#3 was coaxial (0.4 kg) and IHX#4 was core type (0.8 kg).
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 46 of 99

SAE IMAC IHX1: Quality/SH sweep @I35a (full capacity)

1

2

3

4

5

6

7

8

0.8 0.85 0.9 0.95 1 1.05 1.1

Quality [-]

Qe [kW] / COP [-]

COP - baseline

COP - IHX1

COP - IHX2

COP - IHX3

COP - IHX4

Qe - baseline

Qe - IHX1

Qe - IHX2

Qe - IHX3

Qe - IHX4

SAE IMAC IHX1: Quality/SH sweep @M35a (full capacity) 

1

2

3

4

5

6

7

8

0.85 0.9 0.95 1 1.05 1.1

Quality [-]

Qe [kW] / COP [-]

COP - baseline

COP - IHX1

COP - IHX2

COP - IHX3

Qe - baseline

Qe - IHX1

Qe - IHX2

Qe - IHX3
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 47 of 99

SAE IMAC IHX1: Quality/SH sweep @I25a (controlled capacity Ten = 10degC)

1

2

3

4

5

6

7

8

0.85 0.9 0.95 1 1.05 1.1

Quality [-]

Qe [kW] / COP [-]

COP - baseline

COP - IHX1

COP - IHX2

COP - IHX3

COP - IHX4

Qe - baseline

Qe - IHX1

Qe - IHX2

Qe - IHX3

Qe - IHX4

SAE IMAC IHX1: Quality/SH sweep @M25a (controlled capacity Ten =10 degC)

1

2

3

4

5

6

7

8

0.85 0.9 0.95 1 1.05 1.1

Quality [-]

Qe [kW] / COP [-]

COP - baseline

COP - IHX1

COP - IHX2

COP -IHX3

Qe - baseline

Qe - IHX1

Qe - IHX2

Qe - IHX3
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 48 of 99

No real benefit was noted with IHX study two, but this work was done with a more 

effective evaporator and the OEM team determined that this was the major reason that 

minimal benefits were observed in this study. Clearly, the OEM team believes IHX 

benefits depend on the evaporator design.

Insulated Suction Line:

As part of the BoB system test, insulating the suction line was evaluated at select 

screening points and for this system showed no significant improvement as illustrated 

below using the IHX line.

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

I35a without insulation I35a with insulation

condition

COP / Qe

COP Qe
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 49 of 99

Improved Variable Compressors:

Initially, two compressor suppliers submitted improved designs that were tested against 

the baseline compressor. While an improvement was measured, the team felt further 

improvement should be possible. The two suppliers were asked to supply second 

generation design. Both supplied newer designs that were also tested. 

Improved Variable Compressor Study one:

The results shown below are for the first two compressors which were reported out in 

the December, 2005 Interim report.

SAE IMAC system component table for compressor study: system No.1

Served as baseline

Component Feature Comp#1 Comp#2 Comp#3

Condenser (Cond#5)

Core effective face [mm] W626xH386XD16

Face area [mm2]

 241,636

Core volume [mm3] 3,866,176

Aspect ratio [-] 0.62

No. of tubes 39 (micro-channel)

No. of Passes 3 (18-9-12)

Mass [kg] 2.1

Evaporator (Evap#1)

Core effective face [mm] W266xH251xD38

Face area [mm2]

 66,766

Core volume [mm3] 2,537,108

Aspect ratio [-] 0.94

No. of tubes 39 (micro-channel)

No. of Passes two slabs, passes unknown

Mass [kg] 1.2

Compressor

Type External controlled 

variable displacement

External 

controlled 

variable 

displacement

External 

controlled 

variable 

displacement

Displacement [cc] ~160 ~160 ~160

Mass (including pulley) [kg] 6.4 6.1 6.5

Receiver/Drier Mass [kg] IRD, see condenser

Total Mass [kg] 9.7 9.4 9.8

Note:

1. Three compressors were tested with this system.

2. EXV as expansion device
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 50 of 99

SAE IMAC Compressor comparison: Evaparator Capacity Q

Note: Controlled Ten=10degC for L25, M25 and 15degC points 

0

1

2

3

4

5

6

7

8

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

Qe [kW]

Compr #1

Compr #2

Compr #3

SAE IMAC Compressor comparison: COP

0

1

2

3

4

5

6

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

COP [-]

Compr #1

Compr #2

Compr #3
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 51 of 99

SAE IMAC Compressor comparison: Wcomp

0

0.5

1

1.5

2

2.5

3

3.5

4

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

Wcomp [kW]

Compr #1

Compr #2

Compr #3

Annual results for the 1st round of improved compressors are reported below with the 2nd test.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 52 of 99

Improved Variable Compressor Study two:

The results shown below are for the final two compressors. The condenser was 

changed from condenser #5 to condenser #3 for these tests.

SAE IMAC system component table for compressor study: system No.2

Served as baseline

Component Feature Comp#1 Comp#4 Comp#6

Condenser (Cond#3)

Core effective face [mm] W626xH386XD16

Face area [mm2]

 241,636

Core volume [mm3] 3,866,176

Aspect ratio [-] 0.62

No. of tubes 39 (micro-channel)

No. of Passes 3 (22-11-6)

Mass [kg] 2.1

Evaporator (Evap#1)

Core effective face [mm] W266xH251xD38

Face area [mm2]

 66,766

Core volume [mm3] 2,537,108

Aspect ratio [-] 0.94

No. of tubes 39 (micro-channel)

No. of Passes two slabs, passes unknown

Mass [kg] 1.2

Compressor

Type External controlled 

variable displacement

External 

controlled 

variable 

displacement

External 

controlled 

variable 

displacement

Displacement [cc] ~160 ~160 ~160

Mass (including pulley) [kg] 6.4 6.5 (est.) 5.9 (est.)

Receiver/Drier Mass [kg] IRD, see condenser

Total Mass [kg] 9.7 9.8 9.2

Note:

1. Three compressor were tested with this system.

2. EXV as expansion device
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 53 of 99

SAE IMAC Compressor comparison: Evaparator Capacity Qe

Note: (1) Maximum Qe for 45decC 35degC points and 

(2) controlled Ten=10degC for 25degC 15degC points 

0

1

2

3

4

5

6

7

8

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

Qe [kW]

Comp#1 (supplier A IMAC sample I -- Baseline) with cond#3

Comp#4 (supplier B IMAC sample II) with cond#3 

Comp#6 (supplier A IMAC sample III) with cond#3

SAE IMAC Compressor comparison: COP

Note: (1) Maximum Qe for 45decC 35degC points and 

(2) controlled Ten=10degC for 25degC 15degC points 

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

COP [-]

Comp#1 (supplier A IMAC sample I -- Baseline) with cond#3

Comp#4 (supplier B IMAC sample II) with cond#3 

Comp#6 (supplier A IMAC sample III) with cond#3 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 54 of 99

SAE IMAC Compressor comparison: Wcomp

Note: (1) Maximum Qe for 45decC 35degC points and 

(2) controlled Ten=10degC for 25degC 15degC points 

0

0.5

1

1.5

2

2.5

3

3.5

4

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

Wcomp [kW]

Comp#1 (supplier A IMAC sample I -- Baseline) with cond#3

Comp#4 (supplier B IMAC sample II) with cond#3 

Comp#6 (supplier A IMAC sample III) with cond#3 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 55 of 99

Improved Variable Compressor COP and Energy Annualized Weighting:

COP:

Test Point Compr#1 Compr#2 Compr#3 Compr#1 Compr#4 Compr#6

Cond #5 Cond #5 Cond #5 Cond #3 Cond #3 Cond #3

I45 2.07 2.37 2.31 2.31 2.49 2.59

L45 1.92 2.10 2.06 2.01 2.05 2.31

I35a 2.96 3.04 3.03 2.96 3.16 3.41

L35a 2.60 2.65 2.54 2.57 2.51 2.93

M35a 1.88 1.97 1.91 1.91 1.87 2.15

I25a 3.40 3.47 3.29 3.44 3.60 3.92

L25a 3.07 3.63 3.72 3.24 3.60 3.52

M25a 2.48 3.06 3.40 2.66 3.24 2.98

I15 3.22 3.88 3.72 3.49 4.02 3.90

L15 1.90 2.13 1.86 2.17 3.44 2.67

M15 1.20 1.33 1.41 1.29 2.50 1.38

Weighted: 2.07 2.38 2.38 2.22 2.90 2.51

% Improved compared to 

Comp#1/Cond#5

14.90% 14.90% 7.16% 39.77% 21.30%

Energy:

Energy Comparison

Test Point Compr#1 Compr#2 Compr#3 Compr#1 Compr#4 Compr#6

Cond #5 Cond #5 Cond #5 Cond #3 Cond #3 Cond #3

I45 1.763 1.539 1.573 1.580 1.465 1.407

L45 2.585 2.367 2.408 2.468 2.420 2.151

I35a 1.531 1.494 1.495 1.534 1.436 1.331

L35a 2.344 2.298 2.393 2.373 2.427 2.076

M35a 3.626 3.461 3.583 3.570 3.650 3.179

I25a 1.190 1.169 1.232 1.177 1.127 1.034

L25a 0.904 0.765 0.746 0.858 0.771 0.789

M25a 1.097 0.888 0.798 1.021 0.838 0.911

I15 0.190 0.158 0.164 0.175 0.152 0.157

L15 0.234 0.209 0.240 0.205 0.129 0.166

M15 0.337 0.305 0.287 0.314 0.162 0.292

Weighted: 710,187 629,853 624,883 675,810 580,323 603,348

% change compared to 

Comp#1/Cond#5

-11.31% -12.01% -4.84% -18.29% -15.04%
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 56 of 99

Condenser Sub-cooling Pass Design:

At the beginning of the IMAC project, team two solicited condenser samples with 

varying subcooling pass configurations that are illustrated below. These condensers 

were designed to enable use to evaluate the effects of these configurations on system 

efficiency. Five configurations were submitted, but only the extremes were evaluated.

The total number of tubes for each condenser was 39. Each heat exchanger had two 

passes plus sub-cooler except condenser #1 which does not have a sub-cooling pass. 

Number of tubes in each pass is shown in the square. Arrow indicates the flow 

direction. Refrigerant circuiting may not be totally optimized but it is representative of 

these design types. 

Condenser 1

Condenser 3

Condenser 5

18

9

11

22

13

26
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 57 of 99

Three condenser circuiting designs were evaluated. 

SAE IMAC system component table for subcooling study

Served as baseline

Component Feature Cond#1 Cond#3 Cond#5

Condenser

Core effective face [mm] W626xH386XD16 W626xH386XD16 W626xH386XD16

Face area [mm2]

 241,636 241,636 241,636

Core volume [mm3] 3,866,176 3,866,176 3,866,176

Aspect ratio [-] 0.62 0.62 0.62

No. of tubes 39 (micro-channel) 39 (micro-channel) 39 (micro-channel)

No. of Passes 2 (26-13) 3 (22-11-6) 3 (18-9-12)

Mass [kg] 1.8 2.1 2.1

Evaporator (Evap#1)

Core effective face [mm] W266xH251xD38

Face area [mm2]

 66,766

Core volume [mm3] 2,537,108

Aspect ratio [-] 0.94

No. of tubes 39 (micro-channel)

No. of Passes two slabs, passes unknown

Mass [kg] 1.2

Compressor (Comp#1)

Type External controlled variable displacement

Displacement [cc] 160

Mass (including pulley) [kg] 6.4

Receiver/Drier Mass [kg] 0.4 IRD, see condenser IRD, see condenser

Total Mass [kg] 9.8 9.7 9.7

Note:

1. IMAC Cond#1 #3 and #5 are the same in size and design except for pass arrangement.

2. Condenser mass Includes IRD (Integrated Receiver/Drier) if applicable.

3. IMAC Cond#1 #3 and #5 use the same evaporator and the same compressor. 

Table shows percentage of the surface dedicated to sub-cooling.

SAE IMAC Condenser Circuiting

Number of tubes Sub-cooling to 

First pass Second pass Sub-cooling total surface [%]

pass

Condenser #1 26 13 0 0

Condenser #3 22 11 6 18

Condenser #5 18 9 12 44
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 58 of 99

SAE IMAC R134a charge determination test condenser#1, 3, 5: COP vs DTsub

2

2.2

2.4

2.6

2.8

3

3.2

3.4

3.6

3.8

0 5 10 15 20 25 30

DTSub [K]

COP [-]

COP_L35 cond#1 COP_I35a cond#1 COP_I25a cond#1

COP_I35a cond#3 COP_I25a cond#3 COP_L35 cond#5

COP_I35a cond#5 COP_I25a cond#5

COP vs subcooling

SAE IMAC R134a charge determination test condenser#1, 3, 5: Capacity vs DTsub

3

3.5

4

4.5

5

5.5

6

6.5

0 5 10 15 20 25 30

DTSub [K]

Capacity [kW]]

L35 cond#1 I35a cond#1 I25a cond#1 I35a cond#3

I25a cond#3 L35 cond#5 I35a cond#5 I25a cond#5

Capacity vs Subcooling
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 59 of 99

SAE IMAC R134a charge determination test condenser#1, 3, 5: COP vs charge amount

2

2.2

2.4

2.6

2.8

3

3.2

3.4

3.6

3.8

0 200 400 600 800 1000 1200 1400 1600 1800

charge [gm]

COP [-]

COP_L35 cond#1 COP_I35a cond#1 COP_I25a cond#1

COP_I35a cond#3 COP_I25a cond#3 COP_L35 cond#5

COP_I35a cond#5 COP_I25a cond#5

COP vs Charge

SAE IMAC Capacity comparison: cond#1 cond#3 cond#5

0

1

2

3

4

5

6

7

8

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

Q [kW]

Q - Cond#1

Q - Cond#3

Q - Cond#5
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 60 of 99

SAE IMAC COP comparison: cond#1 cond#3 cond#5

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

COP [-]

COP - Cond#1

COP - Cond#3

COP - Cond#5

SAE IMAC Wcomp comparison: cond#1 cond#3 cond#5

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

I45 L45 I35a L35a M35a I25a L25a M25a I15 L15 M15

Test condition

Wcomp [kW]

Wcomp - Cond#1

Wcomp - Cond#3

Wcomp - Cond#5
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 61 of 99

Sub-cooling Control Strategy:

To evaluate the effect of varying sub-cooling with the same condenser, a by-pass 

configuration was devised to allow liquid to separate in the receiver and by-pass the 

sub-cooling loop. Additional valves were added to attempt to take the gas/liquid from 

the top of the receiver and route this to the inlet of the sub-cooling pass to take better 

advantage of the heat exchange capability of this part of the condenser at low loads. 

Below is an illustration of this plumbing. The results were less than satisfactory and will 

be discussed later in the report. It is believed that the liquid separation may not have

been adequate to provide the expected advantage.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 62 of 99

The evaluation of controlled sub-cooling system did not provide a significant 

improvement in COP. However, after review of the data in more detail, the team feels 

that the receiver may not have separated the liquid well during the test and improved 

hardware may produce more promising results. 

SAE IMAC cond#5 bypass subcooling controll: Qe and COP @ I25a

Note: Qe not controlled, compressor run full displacement @ I25a

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

4.5

0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100%

% of flow through subcooling loop

Qe [kW] / COP [-]

Qe - I25a

COP - I25a

Poly. (Qe - I25a)

Poly. (COP - I25a)

COP and Qe results for subcooling control

The results above indicate that there are minimal benefits from the setup and operating 

strategy used. Team 2 feels that there still may be merit but a better method of 

separating the gas from the liquid in the receiver will be necessary to better evaluate 

this technology. 

Since there was not significant improvement noted at these two screening points an 

annual number was not calculated.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 63 of 99

Superheat Control Strategy:

An Electronic Expansion Valve was used. The expansion valve controlled superheat at 

equal capacity and measured the effect on system COP. This study was done first on 

the baseline 38 mm. thick evaporator and then later on the optimized 50 mm thick 

evaporator.

Superheat control study one:

SAE IMAC system component table for superheat study

Component Feature

Condenser (Cond#5)

Core effective face [mm] W626xH386XD16

Face area [mm2]

 241,636

Core volume [mm3] 3,866,176

Aspect ratio [-] 0.62

No. of tubes 39 (micro-channel)

No. of Passes 2 (26-13)

Mass [kg] 2.1

Evaporator (Evap#1)

Core effective face [mm] W266xH251xD38

Face area [mm2]

 66,766

Core volume [mm3] 2,537,108

Aspect ratio [-] 0.94

No. of tubes 39 (micro-channel)

No. of Passes two slabs, passes unknown

Mass [kg] 1.2

Compressor (Comp#2)

Type External controlled variable 

displacement

Displacement [cc] 160

Mass (including pulley) [kg] 6.1

Receiver/Drier Mass [kg] IRD, see condenser

Total Mass [kg] 9.4

Note: Condenser mass Includes IRD (Integrated Receiver/Drier) if 

applicable.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 64 of 99

The superheat was then varied from 2 to 10K. Result can be found in the graph below.

Effect of Superheat on System COP

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

0 1 2 3 4 5 6 7 8 9 10

Superheat at compressor inlet [oC]

COP [-]

L25a

M25a

I15

L15

M15 

L15 Superheat Study Evaporator Grid temperature distribution

0

2

4

6

8

10

12

14

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20

EVAP temp [deg C]

0.5 deg SH

2.4 deg SH

3.2 deg SH

Baseline-7K Baseline-7K 4K SH 4K SH 1K SH 1K SH
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 65 of 99

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

kJ kJ

625,497 2.31 548,603 2.20 522,346 2.96 Energy Comparison

-12.3% -4.4% -16.5% 28.6% COP comparison:

Superheat control study two:

SAE IMAC system component table for SH study #2

Component Feature B-o-B - without IHX 

Condenser (Cond#7)

Core effective face [mm] W546xH396xD22

Face area [mm2]

 216,216

Core volume [mm3] 4,756,752

Aspect ratio [-] 0.73

No. of tubes 61 (micro-channel)

No. of Passes 4 (29-18-8-6)

Mass [kg] 2.5

Evaporator (Evap#3)

Core effective face [mm] W267xH257xD50

Face area [mm2]

 68,619

Core volume [mm3] 3,430,950

Aspect ratio [-] 0.96

No. of tubes micro-channel

No. of Passes two slabs, passes unknown

Mass [kg] 1.6 (est.)

Compressor (Comp#4)

Type External controlled variable displacement

Displacement [cc] ~160

Mass (including pulley) [kg] 6.4 (est.)

Receiver/Drier Mass [kg] IRD, see condenser

IHX Mass [kg] N/A

Total Mass [kg] 9.4
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 66 of 99

SAE IMAC Superheat study at L25a condition

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

4.5

5.0

0 2 4 6 8 10 12 14 16

DTsup [K]

Qe [kW] / COP [-]

Qe - Evap#3

COP - Evap#3

SAE IMAC Superheat study at M25a condition

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

4.5

0 2 4 6 8 10 12 14 16 18

DTsup [K]

Qe [kW] / COP [-]

Qe - Evap#3

COP - Evap#3
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 67 of 99

SAE IMAC Superheat study at L15 condition

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

0 1 2 3 4 5 6 7 8 9

DTsup [K]

Qe [kW] / COP [-]

Qe - Evap#3

COP - Evap#3

SAE IMAC Superheat study at L15 condition

0.0

0.5

1.0

1.5

2.0

2.5

0 1 2 3 4 5 6 7 8

DTsup [K]

Qe [kW] / COP [-]

Qe - Evap#3 

COP - Evap#3 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 68 of 99

Best of the Best System:

The Baseline for comparison is the R134a system from ARCRPII that was run to normal 

evaporator air off temperatures.

The Best of the Best (B-o-B) system includes:

 B-o-B Improved Condenser

 B-o-B Improved Evaporator

 B-o-B Improved External Control Variable Compressor 

 Superheat Control Algorithm w/ EXV

The B-o-B System was run to match ARCRPII evaporator air off temperatures and also 

at elevated temperatures.

SAE IMAC system component table for Best-of-Best System

Component Feature B-o-B - without IHX 

Condenser (Cond#7)

Core effective face [mm] W546xH396xD22

Face area [mm2]

 216,216

Core volume [mm3] 4,756,752

Aspect ratio [-] 0.73

No. of tubes 61 (micro-channel)

No. of Passes 4 (29-18-8-6)

Mass [kg] 2.5

Evaporator (Evap#3)

Core effective face [mm] W267xH257xD50

Face area [mm2]

 68,619

Core volume [mm3] 3,430,950

Aspect ratio [-] 0.96

No. of tubes micro-channel

No. of Passes two slabs, passes unknown

Mass [kg] 1.6 (est.)

Compressor 

(Comp#4)

Type External controlled variable displacement

Displacement [cc] ~160

Mass (including pulley) [kg] 6.4 (est.)

Receiver/Drier Mass [kg] IRD, see condenser

IHX Mass [kg] N/A

Total Mass [kg] 9.4
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 69 of 99

SAE IMAC compared to Phase II: Evaparator Capacity Qe

Note: Maximum Qe for 45/35degC and controlled Qe for 25/15degC 

0

1

2

3

4

5

6

7

8

I60a

I60b

I60c

I45

L45

M45

H45

I50aI35a

L35a

M35a

H35a

I35b

L35b

M35b

H35b

I40aI25a

L25a

M25a

H25a

I40bI25b

L25b

M25b

H25b

I25c

L25c

M25c

H25c

I30I15

L15

M15

H15

Test condition

Qe [kW]

Phase II ARCRP

IMAC B-O-B match phase II

IMAC B-O-B 10 deg C

SAE IMAC compared to Phase II: Evaparator Capacity Qe

Note: Maximum Qe for 45/35degC and controlled Qe for 25/15degC 

0

1

2

3

4

5

6

I60a

I60b

I60c

I45

L45

M45

H45

I50aI35a

L35a

M35a

H35a

I35b

L35b

M35b

H35b

I40aI25a

L25a

M25a

H25a

I40bI25b

L25b

M25b

H25b

I25c

L25c

M25c

H25c

I30I15

L15

M15

H15

Test condition

COP [-]

Phase II ARCRP

IMAC B-O-B match phase II

IMAC B-O-B 10 deg C
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 70 of 99

SAE IMAC compared to Phase II: Wcomp

Note: Maximum Qe for 45/35degC and controlled Qe for 25/15degC 

0

1

2

3

4

5

6

I60aI60c

L45

H45

I35a

M35a

I35b

M35b

I40a

L25a

H25a

I25b

M25b

I25c

M25c

I30

L15

H15

Test condition

Wcomp [kW]

Phase II ARCRP

IMAC B-o-B match phase II

IMAC B-o-B 10 C

The numbers have been annualized a number of ways. First, as compared to the 

ARCRP Phase II R134a system and then as compared to the original baseline system 

that was assembled for the IMAC project. For the comparison below, the data was 

analyzed with equal evaporator air outlet temperatures and the capacity was averaged 

for the energy calculation to assure comparable results. Then the tests were rerun with 

the evaporator out air temperature set to 10°C and the comparison in results is shown 

below.

ARCRP Base IMAC B-o-B

Teao= ARCRP

IMAC B-o-B

Teao= 10C

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

Compr 

Energy

COP 

weighted

610,230 2.86 535,406 3.29 388,115 3.22

-12.3% 15.0% -36.4% 12.6%
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 71 of 99

The IMAC baseline did not run the complete test matrix, so this analysis is done with a 

reduced test matrix. [I45, L45, I35a, L35a, M35a, I25a, L25a, M25a, I15, L15, M15] You 

can see that the IMAC baseline was less efficient than the original ARCRP reference. 

But the improvement in COP is significant as compared to the previous comparison. 

The improvement in energy consumption is slightly smaller, but all of these tests were 

run with the evaporator air outlet temperature equal to 10°C.

IMAC Base 

Compr energy

IMAC B-o-B 

Compr energy

IMAC Base 

COP

IMAC B-o-B 

COP

663,495 481,791 2.20 3.39

27.4% 54.3%
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 72 of 99

Fixed vs Variable Compressors:

The OEM members selected two production fixed displacement compressors for test. 

One is a so-called High Efficiency Piston design and the other is a Scroll design. The 

manufacturers were asked to supply sample compressors. Both suppliers provided 

compressors and assistance specifying oil type and circulation rates. The baseline 

compressor for this test is the BoB variable. The OEM team believed that a cycling 

system would behave differently with an EXV compared to a TXV. It was decided to 

conduct this testing with a TXV. The TXV was specified by the evaporator supplier. 

The baseline variable compressor was run first with the TXV.

SAE IMAC system component table for fixed compressor study

Served as baseline

Component Feature comp#4 comp#7 comp#8

Condenser (Cond#7)

Core effective face [mm] W546xH396xD22

Face area [mm2]

 216,216

Core volume [mm3] 4,756,752

Aspect ratio [-] 0.73

No. of tubes 61 (micro-channel)

No. of Passes 4 (29-18-8-6)

Mass [kg] 2.5

Evaporator (Evap#3)

Core effective face [mm] W267xH257xD50

Face area [mm2]

 68,619

Core volume [mm3] 3,430,950

Aspect ratio [-] 0.96

No. of tubes micro-channel

No. of Passes two slabs, passes unknown

Mass [kg] 1.6 (est.)

Compressor

Type External controlled 

variable displacement

Fixed displacement 

scroll type

Fixed displacement 

piston type

Displacement [cc] ~160 ~90 ~160

Mass (including pulley) [kg] 6.5 (est.) 4.5 5.6

Receiver/Drier Mass [kg] IRD, see condenser

Total Mass [kg] 10.6 8.6 9.7

Note:

1. Three compressors were tested with this system.

2. TXV as expansion device
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 73 of 99

SAE IMAC Variable vs. Fixed compressor: Evaparator Capacity Qe

0

1

2

3

4

5

6

7

8

I45

L45

H45

I35a

L35a

M35a

H35a

I25a

L25a

M25a

H25a

I15

L15

M15

H15

Test condition

Qe [kW]

IMAC TXV data (used as fixed compressor

baseline)

IMAC comp#8 (fixed) data

IMAC comp#7 (fixed) data

SAE IMAC Variable vs. Fixed compressor: COP

0

1

2

3

4

5

6

7

I45 L45 H45 I35a L35a M35a H35a I25a L25a M25a H25a I15 L15 M15 H15

Test condition

COP [-]

IMAC TXV data (used as fixed compressor

baseline)

IMAC comp#8 (fixed) data

IMAC comp#7 (fixed) data
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 74 of 99

SAE IMAC Variable vs. Fixed compressor: Wcomp

0

1

2

3

4

5

6

I45 L45 H45 I35a L35a M35a H35a I25a L25a M25a H25a I15 L15 M15 H15

Test condition

Wcomp [kW]

IMAC TXV data (used as fixed compressor

baseline)

IMAC comp#8 (fixed) data

IMAC comp#7 (fixed) data

Annualization:

IMAC

EXV

IMAC

TXV

Fixed

Compr 

#7

Fixed 

Compr 

#8

IMAC

EXV

IMAC

TXV

Fixed 

Compr 

#7

Fixed 

Compr 

#8

Compr 

Energy

Compr 

Energy

Compr 

Energy

Compr 

Energy

COP 

weighted

COP 

weighted

COP 

weighted

COP 

weighted

kJ

483,575 482,277 576,543 613,494 3.35 3.47 3.08 2.82

0% base 20% 27% -3% base -11% -19%
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 75 of 99

III. Vehicle Testing

A Cadillac STS was used as a vehicle test bed to demonstrate that an improved 

efficiency refrigerant system would provide equal cooling capacity. The plan also 

included development of control strategy algorithms. 

Demonstration of equal capacity:

The main objective of the test vehicle was to demonstrate equal AC performance. This 

was proved during the Phoenix Alternative Refrigerant System Symposium (shown 

below) during the ride/drive evaluation (varying city traffic / Idle driving pattern). 

Panel Outlet Temperatures

ARSS 2006

49.7

54

57.7

55.8

56.8

55.7

58.5

37.3

43.7

46.3

43.3 43.8

45.3

46.8

35

40

45

50

55

60

65

Visteon Audi

R744

Hyundai - H Hyundai -

Prod

Honeywell 

Jetta

I-MAC / Load

reduction

I-MAC /

Optimized Ref. 

I-MAC /

Production. 

Degrees F.

City Traffic Road
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 76 of 99

Vehicle Comfort 2006 ARSS

Vehicle G I-MAC / Production

0

1

2

3

4

5

6

City 1 City 2 City 3 City 4 30 min 35 min 40 min 45 min

Comfort Rating

Driver Rt. Front Lft. Rear Rt. Rear

Comfort

All comfort ratings based upon average of occupied seats

Vehicle Comfort 2006 ARSS

Vehicle F I-MAC / Optimized Ref.

0

1

2

3

4

5

6

City 1 City 2 City 3 City 4 30 min 35 min 40 min 45 min

Comfort Rating

Driver Rt. Front Lft. Rear Rt. Rear

Comfort

Team 7, 10 and 12 rides

System problems for 8 and 11 rides 

All comfort ratings based upon average of occupied seats

2 of 3 evaluators 

never reach a 

5-comfort (best)

2 of 3 evaluators reach a 

5-comfort (best), better 

than production 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 77 of 99

Shown below is a city traffic comparison between Team 3’s Load reduction vehicle with a 

production STS AC refrigerant system and the Team 2 Efficiency vehicle, showing equal 

performance. 

Average Vent Outlet Temperature Comparison Showing Equal Performance Between Team 2 

Efficiency Vehicle and Thermal Load Reduction Vehicles

( Ambient 43 C,100 % Outside, 650 W/m^2 average solar, June 21, 2006 Ariz.)

-5

0

5

10

15

20

25

30

35

40

100 105 110 115 120 125 130 135 140

Time (s)

Average Vent Outlet Temperature (°C), Recirc 0 %

Efficiency Load Reduction (CAN) AirInlet Door Pot %

Efficency

Vehicle

Thermal Load 

Reduction Vehicle
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 78 of 99

The vehicle development plan is shown below which progressed with 3 generations of 

algorithm improvement, with a final verification of the 3rd/ final generation.

Vehicle Build and Test Plan

Generation One:

The vehicle was built with an optimized external controlled variable compressor, an 

electronic expansion valve, D-Space controller, MATLAB and SIMULINK software which 

used a new algorithm for control of the compressor and expansion valve to optimized 

COP control. All other components were original equipment parts from the 2005 

Cadillac STS.

The first generation of software was produced by EDS from GM based software. The 

algorithm was designed to operate in two phases. The first was a pull down region 

where the cabin temperature is greater than the desired temperature. In this region the 

compressor displacement was controlled to the maximum displacement, the expansion 

valve was controlled to maintain optimal superheat for maximum capacity, and the air 

recirculation door was set to 0% outside air. The second phase, (comfort region) took 

over when the desired cabin temperature has been reached, and the capacity could be 

reduced to minimize energy usage, while maintaining occupant comfort. In the comfort 

region, the compressor displacement is reduced as much as possible to raise the

evaporator outlet pressure keeping the evaporator temperature as high as possible 

consistent with vehicle occupant comfort. The recirculation door is opened periodically 

to allow for sufficient outside air for occupant comfort, and the EXV is controlled to 

maintain an optimal superheat for maximum COP.

Task Responsibility Completion Comments

Obtain Vehicle Bill Hill

Complete - see 

Note 1

GM will provide 2 Cadillac STS Vehicles 

(Efficiency and Heatload)

Identify Test Procedures: Team 2 OEMs

Select Refg Control Parameters Team 2 Complete Based on stand test at University of Illinois

Data Acquisition Bill Hill Complete GM will provide 2 systems

Source Refg Sys Controls Alg Dev Bill Hill/Supplier Complete EDS

Parts to Nissan Suppliers/University of Illinois Complete Maintain supplier confidentiality

Install EXV + Data Acquisition GM Complete Install EXV, datalogger, dSpace Controller

Prep & Instrument Vehicles Dave Barwin Complete Nissan to install parts and instrument

Vehicle Algorithm dev Team 2 Complete In Detroit

Ship Vehicles to Golden, CO Bill Hill Complete

Vehicle Drive Eval - Golden, CO Bill Hill and Dave Barwin Complete

Vehicle at NREL - Golden, CO Bill Hill Complete Soak & cool down test for CAE Model Corr

Vehicle Algorithm dev Team 2 Complete In Detroit

Wind Tunnel Test Team 2 Sept 20-21, 2005 Scrubbed due to instrumentation problems

Develop Future Vehicle Test Plans Team 2 Complete

Dedicated development engineer needed to 

continue the vehicle test plan

Algorithm contract with CTS CTS / Dave Barwin Complete Outsourced development to CTS

Test Drive OEMs / Dave Barwin Complete

Install Supplier B compressor Supplier B Complete

Additional Algorithm development CTS / Dave Barwin Complete

Ship Vehicles to Phoenix Complete

Prep Drive in Phoenix OEMs / Dave Barwin Complete

Sponsor Ride Complete

Phoenix Symposium OEMs / Dave Barwin Complete

Ship Vehicles to NREL Complete Solar soak comparison to Team 3

Ship Vehicle to Detroit Complete

Phase 3 verify Super heat vs Cop CTS / Dave Barwin Complete Completed

Team 2 Efficiency Vehicle Test Plan-Plan C

Objective: Demonstrate that an improved efficiency A/C system meets normal cooldown performance.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 79 of 99

The vehicle was fully instrumented and outfitted with several components to achieve the 

desired control and measure the effects. A D-Space MicroAutobox is used along with a 

driver-box to read the parameters valid for control, run the control algorithm, and output 

the desired control signals to the compressor and EXV controller. The following 

diagram shows the vehicle hardware setup.

The software model was implemented in MatLab Simulink using the dSpace RTI library 

and executes on the MicroAutobox controller. The software model is divided into four 

layers, the Hardware Input Output (HWIO) Layer which reads the data inputs and 

provides the controller outputs, the Software Integration (SWIL) Layer which handles 

general services, the Application Layer which composes the main body of the 

application, and the Operating System (OS) Layer which provides controller 

management.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 80 of 99

The software model includes a real time calculation of COP and Capacity. Various 

methods have been explored to estimate these values. The necessary parameters to

calculate these values are all recorded excluding compressor power. The compressor 

power was calculated in generation one originally using a normalized load calculation 

and the refrigerant flow rate, which was calculated by multiplying compressor 

displacement (Disp), duty cycle, speed (Vc), suction density (ρs) and volumetric 

efficiency (ηv) as shown in the equation below. The duty cycle was estimated based on 

the control signal to the compressor and the volumetric efficiency was estimated based 

on discharge pressure. Using the mass flow rate from this method predicted 

unreasonably high capacities originally.

It was also determined that the original expansion control valve driver was too slow and

that the hardware/sensors needed to be calibrated. The data acquisition system (DAS) 

was also overheating which need to be resolved. 

Creative Thermal Solutions (CTS) in Urbana, IL was chosen to debug the vehicle and 

algorithm to provide additional functionality to the A/C system control and improve the 

algorithm. 

CTS developed a new electronic expansion valve driver control board and software, to 

be able to speed up the superheat adjustment, they resolved the DAS overheating, 

calibrated the pressure transducers, made corrections to the COP and normalized load 

calculation, and re-wired the Micro Auto box inputs. 

To evaluate these changes the vehicle was taken on a test drive, however after some 

corrections to the implementation of this mass flow rate (M r) equation the prediction 

preformed better as is shown in the capacity plot as ‘EDS method corrected’.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 81 of 99

Vehicle Tests vs. IMAC Bench Data - Capacity Comparison for EDS and Corrected EDS 

Methods

0

2

4

6

8

10

12

14

16

I35b L35b I25b L25b

Capacity ( kW )

Vehicle EDS equations Vehicle EDS equations corrected Capacity bench data

 .

The corrected calculation (found units errors and multiplying by specific volume instead of 

dividing by it) gave reasonable capacity values at the full displacement I25b point but 

performed poorly at reduced displacement points because there was no accounting for 

the loss in volumetric efficiency at reduced displacements and because the duty cycle 

did not directly correlate to the displacement. The torque was also estimated using a 

normalized load calculation which resulted in high predictions of compressor work.

It was determined during the test drive from Urbana, IL to Laredo, TX., that the 

algorithm still needed further improvement to calculate COP and capacity. The link 

between efficiency and displacement was difficult to separate and estimate with 

reasonable accuracy. 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 82 of 99

The following plot shows the effect of changes on compressor speed and evaporator air 

exit temperature (EAT) on COP. While the algorithm predicted COP values in a 

reasonable range for part of the test, clearly the prediction was not reasonable over the 

entire test. 

Texas Ride Data - Effect of Compressor Speed and EAT on COP

0

1

2

3

4

5

6

7

8

9

10

0 1000 2000 3000 4000 5000 6000

Time (s)

COP ( - ), Temperature (C)

0

500

1000

1500

2000

2500

3000

3500

4000

4500

Compressor Speed (rpm)

COP EAT RPM

Generation two:

Following the Texas ride, the vehicle was returned to Detroit to switch compressors to 

the compressor that performed best at the University of Illinois ACRC labs. At this time 

the method (2nd Algorithm version) for estimating mass flow rate and compressor power 

was changed to use linear regression. The regression was based off of the air flow rate 

on the condenser (AFRc), the air temperature entering the condenser (Tcai), the air 

temperature exiting the evaporator measured at the nozzle (Ten), the compressor speed 

(Vc), the discharge pressure (Pcpro), the compressor displacement control current 

(CompCurr) and the air flow rate over the evaporator (AFRe). 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 83 of 99

The regression constants were fit using data taken at the ACRC. The equations are 

listed below.

A regression for compressor power was done in the same fashion calculating a different 

term (Kp) using the equation below.

This model was found to match a limited number of data points that it was based on 

reasonably well as can be seen in the following plots showing bench data vs. regression 

predictions for refrigerant mass flow rate and compressor power. However, when the 

regressions were used in the vehicle at conditions differing from the bench conditions 

the mass flow rate and compressor work predicted values 2 to 3 times higher than 

expected. 

Mass Flow Rate Estimation

R

2

 = 0.9542

0

10

20

30

40

50

60

0 10 20 30 40 50 60

IMAC Bench Data

Calculated Value

5 6 7

0 1 2 3 4

a CompCurr a AFR a

M a AFR a T a T a V a P

e

r c cai en c cpro

    

         

5 6 7

0 1 2 3 4

b CompCurr b AFR b

P b AFR b T b T b V b P

e

comp c cai en c cpro

    

         
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 84 of 99

Compressor Power Estimation

R

2

 = 0.9836

0

0.5

1

1.5

2

2.5

3

0 0.5 1 1.5 2 2.5 3

IMAC Bench Data

Calculated Value

The vehicle was taken to Phoenix in the summer, 2006 and high temperature testing 

was performed. Interestingly, the effect of superheat on capacity and COP was found 

to be different than had been previously observed on the bench. During idle, the 

optimal superheat for maximum COP and capacity was observed to be 10°C to 15°C 

whereas bench data has indicated that superheats near 0°C are optimal. In order to 

further understand the effect of superheat on COP and capacity, vehicle testing was 

performed in the climatic chamber facility to eliminate dynamics that may have affected 

these results.

Effect of Superheat on COP and Capacity - Phoenix Idle Condition

0

5

10

15

20

25

0 100 200 300 400 500 600 700 800

Time (s)

Temperature (deg C)

0

0.5

1

1.5

2

2.5

3

COP ( - ), Capacity (kW)

Superheat Setpoint Actual Superheat Evaporator Capacity COP

Evaporator Capacity

COP

Actual Superheat

Superheat Setpoint
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 85 of 99

Generation three:

After the Phoenix Alternative Refrigerant System Symposium, the Algorithm was then 

modified (3rd Version), to incorporate a physically based approach. In this approach the 

original mass flow rate equation which calculated mass flow rate by multiplying 

compressor displacement, speed, suction density and volumetric efficiency was used. 

The compressor displacement and volumetric efficiency were combined into one term 

(Kv) and the regression was performed basing this term on a combination of pressure 

ratio (Pd/Ps), compressor pressure difference *(Pd-Ps), compressor speed (Vc), and 

compressor displacement control signal current (CompCurr). The regression equation

(physically based regression) shown below was used with the full ACRC test matrix 

data to calculate the constants C1 through C7.

Kv= (1-C0*Pd/Ps)*(C1*(Pd-Ps)+C2*CompCurr+C3*Vc)+C4+C5*Pd+C6*Ps+C7*CompCurr

A regression for compressor power was done in the same fashion calculating a different 

term (Kp) using the equation below. 

Kp= (1-D0*Pd/Ps)*(D1*(Pd-Ps)+D2*CompCurr+D3*Vc)+D4+D5*Pd+D6*Ps+D7*CompCurr

The results from this regression fit the ACRC test bench data very well as can be seen 

in the following plots showing the regression vs. bench data for COP, capacity, and 

compressor power. It should be noted that the regression predicts high values of COP.

Regression Results for COP vs. IMAC Bench Data

0

1

2

3

4

5

6

7

8

9

10

I60aI60b

I60c

I45

L45

M45I50aI35a

L35a

M35a

I35b

L35b

M35b

I40aI25a

L25a

M25a

I40bI25b

L25b

M25b

I25c

L25c

M25c

COP ( - )

COP - bench COP - vehicle regression

COP values for low compressor displacements as can be seen in the conditions 

M25a, L25b, M25b, L25c, and M25c.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 86 of 99

Regression Results for Capacity vs. IMAC Bench Data

0

1

2

3

4

5

6

7

8

I60aI60b

I60c

I45

L45

M45I50aI35a

L35a

M35a

I35b

L35b

M35b

I40aI25a

L25a

M25a

I40bI25b

L25b

M25b

I25c

L25c

M25c

Capacity (kW)

Capacity - bench Capacity - vehicle regression

Regression Results for Compressor Power vs. IMAC Bench Data

0

0.5

1

1.5

2

2.5

3

3.5

4

I60aI60b

I60c

I45

L45

M45I50aI35a

L35a

M35a

I35b

L35b

M35b

I40aI25a

L25a

M25a

I40bI25b

L25b

M25b

I25c

L25c

M25c

Compressor Power (kW)

Compressor power - bench Compressor power - vehicle regression

The vehicle was tested in the climatic chamber facility at CTS to match 4 conditions 

from the bench test matrix (conditions selected I35b, L35b, I25b, and L25b). The 

results matched the bench data very well showing good performance from the 

regression. The following plots show the vehicle results compared to the bench data for 

COP, capacity and compressor power. It should be noted that for L35b the compressor 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 87 of 99

displacement was higher than on the bench resulting in a higher capacity and 

compressor power.

CTS Vehicle Tests vs. IMAC Bench Data - COP Comparison

0

1

2

3

4

5

6

7

I35b L35b I25b L25b

COP ( - )

COP bench data COP car regression

CTS Vehicle Tests vs. IMAC Bench Data - Capacity Comparison

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

I35b L35b I25b L25b

Capacity ( kW )

Capacity bench data Capacity car regression
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 88 of 99

CTS Vehicle Tests vs. IMAC Bench Data - Compressor Power Comparison

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

I35b L35b I25b L25b

Work (kW)

Wcomp bench data Wcomp car estimated
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 89 of 99

After returning from Phoenix it was decided to study the effect of superheat, and its 

relationship to COP, and Capacity in the vehicle to determine why the 2nd Algorithm 

version gave conflicting results from Bench testing. The study also examined the 

effects of re-circulated air control and evaporator temperature control on compressor 

work. The test points in the vehicle were at IMAC conditions I25b, L25b, and M25b as 

well as a condition matching the Phoenix point (45C, 1000 engine rpm). At each 

condition, the vehicle was tested at superheats of 2°C, 5°C, 10°C, 15°C, and 20°C while 

all other parameters were maintained constant. There was difficulty maintaining the 

superheat constant at the low displacement conditions, particularly at L25b and M25b. 

System instability was observed to be the most severe at superheats of 10°C and 15°C. 

The following plot shows the superheat instability at various superheat set points which

could be eliminated given more time to calibrate the expansion valve Proportional 

Integral Derivative (PID) constants. 

System Instability at Various Superheat Setpoints

0

5

10

15

20

25

0 10 20 30 40 50 60

Time (min)

Superheat (°C)

Measured Superheat Superheat Setpoint
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 90 of 99

Shown below is an example of the Proportional and Derivative constant tuning to give 

better control of superheat.

A refrigerant mass flow meter was installed prior to running the superheat study so the 

capacity could be calculated using this flow rate multiplied by the refrigerant enthalpy 

change across the evaporator. 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 91 of 99

Effect of superheat on Evaporator Capacity at constant Compressor PWM setting:

The following plot shows that for all of the operating conditions tested the capacity drops 

as superheat increases and that a superheat of 2°C provide an optimum capacity. 

During all the superheat variation tests for each condition, the compressor control signal 

was set to a constant value and only the expansion valve setting was varied. 

Effect of Superheat on Capacity

1.5

2

2.5

3

3.5

4

0 5 10 15 20 25

Superheat (°C)

Capacity (kW)

I25 L25 M25 Phoenix
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 92 of 99

Effect of evaporator Superheat on COP at constant capacity:

COP now behaves similar to the test bench. The I25b and Phoenix operating 

conditions were at full displacement and showed that COP was highest at the lowest 

superheats. At the L25b and M25b conditions the compressor displacement was 

adjusted to maintain a constant capacity (2.4kW). 

Effect of Superheat on COP when Q is Fixed

0

0.5

1

1.5

2

2.5

3

3.5

0 2 4 6 8 10 12 14 16 18

Superheat (°C)

COP ( - )

I25 L25 M25 Phoenix
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 93 of 99

It was determined using the EAT sensor gave misleading airside capacity calculation 

results. This is show below at a constant capacity test (2.4kW), while superheat is 

decreased the EAT is increased. Considering EAT alone does not account for latent 

capacity which can be substantial. 

-4

-3

-2

-1

0

1

2

3

4

5

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16

DTsup 1.5 DTsup 9 DTsup 17

At Dtsup = 1.5°C, EAT = 6°C, Grid avg = 1.6°C, Tsat_evap = 0°C, Tevap ref out = 1.5°C

At Dtsup = 9°C, EAT = 5°C, Grid avg = 0°C, Tsat_evap = -1.8°C, Tevap ref out = 7.5°C

At Dtsup = 17°C, EAT = 3.7°C, Grid avg = -0.9°C, Tsat_evap = -4°C, Tevap ref out = 15°C
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 94 of 99

Effect of varied evaporator air set point:

An optimal superheat of 2°C was chosen for the four operating conditions, and EAT was 

varied by reducing compressor displacement to simulate capacity reduction potential 

once the cabin of the vehicle was cooled. The EAT was reduced from approximately 

3°C to 14°C for the conditions with the starting EAT depending on the lowest that could 

be achieved at a given condition. For the Phoenix condition the lowest EAT attainable 

was higher than 14°C so the condition was left out of the EAT study. The following plots 

show the capacity, compressor power reduction, and COP increases attainable by 

allowing the EAT to increase. The COP increases would be less then the estimated 

COP’s show here due to the fact that COP is predicted somewhat high at low 

displacements as shown previously. Significant compressor power savings can be 

achieved by reducing displacement after the cabin of the vehicle is cooled and 

Evaporator set-point logic is invoked.

Capacity and Compressor Power Reduction by Decreasing Displacement and Increasing EAT

0.00

0.50

1.00

1.50

2.00

2.50

3.00

3.50

0 2 4 6 8 10 12 14 16

Evaproator Exit Air Temperature (°C)

Capacity, Compressor Power (kW)

I25 Capacity L25 Capacity M25 Capacity I25 Compressor Power L25 Compressor Power M25 Compressor Power

37.4% Reduction

 M25 Compressor (kW) 

46.7 % Reduction 

 Compressor, I25 (kW)

48% Capacity 

Reduction I25 (kW)

 34.7 % Reduction 

L25 Compressor (kW)

53.2% Capacity

Reduction L25 (kW)
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 95 of 99

Effect of varied re-circulation control on Evaporator Capacity and Compressor Power:

A study to determine the power savings attainable by using higher amounts of 

recirculation air was also performed. For each of the conditions a range of recirculation 

air percentage from 0% to 100% was tested. The following plots show significant 

compressor power and capacity savings can be attained by increasing the air 

recirculation amounts. Furthermore at the I25b and Phoenix conditions the compressor 

was operating at full displacement thus much lower evaporation temperatures were 

attainable at full recirculation.

Capacity Reduction Possible by

 Increasing Air Recirculation Percentage

2.00

3.00

4.00

5.00

6.00

7.00

8.00

9.00

10.00

0 10 20 30 40 50 60 70 80 90 100

Air Recirculation Percentage (%)

Capacity (kW)

I25 L25 M25 Phoenix

L25

M25

55 % Reduction 

Potential I25

 55% Capacity

 Reduction ,Phoenix

Evap Temp

 13.5 C

Evap Temp

30 C

Evap Temp 

17C

Evap 6 C

Compressor Power Reduction Possible by Increasing Air Recirculation Percentage

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

0 10 20 30 40 50 60 70 80 90 100

Air Recirculation Percentage (%)

Compressor Power (kW)

I25 L25 M25 Phoenix
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 96 of 99

Verification of capacity regression [Measured vs Algorithm]:

Adding the refrigerant flow meter allowed for an accurate measurement of system 

capacity and also allowed for a quality check of the physical based regression

algorithm. The following plot shows a comparison of the capacity calculated using the 

mass flow meter and enthalpy change across the evaporator, and the capacity 

calculated using the final physically based regression. All of the data taken in the 

superheat, Evaporator air Temperature [EAT], and recirculation air studies is shown in 

the plot. Compared to the original capacity estimation methods where the capacity was 

predicted 2 to 3 times higher than actual this regression performs very well. 

Regression Capacity Compared to Measuerd Capacity Calculated Using Refrigernat Mass 

Flow Meter

0

1

2

3

4

5

6

7

8

1 3 5 7 9 11 13 15 17 19 21 23 25 27 29 31 33 35 37 39 41 43 45 47

Capacity (kW)

Q regression Q measured
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 97 of 99

Summary of the algorithm comparisons:

The plots below show results from the 3 algorithm iterations calculated with vehicle data 

for the estimation of COP, capacity, and compressor power, as well as bench test 

results at the same conditions, and one additional method for COP calculation (Enthalpy 

based), and another additional method for capacity calculation (Air Side Capacity). This 

should help understand the progression / refinement of the algorithm. The first iteration 

was a physically based approach implemented by EDS, the second was a linear 

regression approach implemented, and the third was a physically based regression

approach. 

CTS Vehicle Tests vs. IMAC Bench Data - COP Comparison

0

1

2

3

4

5

6

7

8

I35b L35b I25b L25b

COP ( - )

COP vehicle EDS equations COP vehicle original regression COP vehicle final regression COP bench data COP vehicle enthalpy based

Enthalpy Based COP Calculation

An additional method for COP calculation was implemented using only the refrigerant 

enthalpies. The COP was calculated by dividing the change in refrigerant enthalpy 

across the evaporator by the change in refrigerant enthalpy across the compressor as 

shown in the equation below. A constant, E1, was calculated using the bench data to 

account for compressor efficiency. 
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 98 of 99

CTS Vehicle Tests vs. IMAC Bench Data - Capacity Comparison

-4

-2

0

2

4

6

8

10

12

14

16

I35b L35b I25b L25b

Capacity ( kW )

Vehicle EDS equations Vehicle original regression Vehicle final regression Capacity bench data Vehicle air side dry

Airside Capacity Calculation

A dry air side capacity measurement was also evaluated. As expected, this method 

outputs low values at the four conditions shown in the capacity plot below due to the 

lack of the latent load component. To calculate the capacities the air flow rate (AFRc)

was multiplied by the specific heat and the change in air temperature across the 

evaporator as shown in the equation below.
IMAC Team 2 - Improved Efficiency Final Report April, 2007

Page 99 of 99

CTS Vehicle Tests vs. IMAC Bench Data - Compressor Power Comparison

-0.5

0

0.5

1

1.5

2

2.5

3

3.5

4

I35b L35b I25b L25b

Power (kW)

Wcomp vehicle EDS equations Wcomp vehicle original regression Wcomp vehicle final regression Wcomp bench data

Conclusion:

The final physically based regression showed a very good agreement with the bench as 

did the enthalpy based COP calculation. Promising results have been observed 

quantifying the potential savings in energy usage by using the control algorithm. By 

reducing the compressor displacement and increasing the EAT after the cabin has been 

cooled, significant energy savings can be achieved. Additional savings can be achieved 

by increasing the air recirculation and reducing the superheat. The control algorithm 

has been set to control all of these parameters to maximize energy savings.