### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 control s 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 wa s to demonstrate an improvement of 30% in system CO P 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 Im proved Mobile Air Conditioning ( IMAC ) Cooperative Research Program (CRP) began in August, 2004. The efficiency team's technical approach started with “brain storming” ideas for potential technologies that would improve efficiency. This list was divided among the members and each member was asked to research individual techno logies 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. Base d on this assessment, the items were prioritized, an d potential suppliers of these technologies were requested to submit sample components. This report presents th e results of the evaluation s that ha ve been conducted on a System Test Stand in the lab and a demonstration vehicle . The project began with org anizational 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 Fin al Report Ap ril, 2007

Page 2 of 99 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 cond

itioning system on a test bench IMAC2-J J2766 Standard for Annualized Climate Calculation of System Power

Consumption of a mobile air conditioning system IMA C2-K User’

s manual for the vehicle controls Summary

of Results and Conclusions: T he 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 3 6 % reducti on in energy consumption as compared to the ARCRP baseline which was run at lower evaporator out temperatures . The COP improvement was 1 2.6 % when compared at this same condition. At equal capacity, the COP improvement was 15%. [The improvement is greate r 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.

AR CRP Base IMAC B -o - B Teao= A RCRP 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 Fin al Report Ap ril, 2007

Page 3 of 99 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 configu red by a new supplier to explore different subcooling pass configuration s . 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 nea rly 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.

IM AC Base Compr energy IM AC B - o - B

Compr energy IM AC Base COP IM AC B - o - B

COP 663, 4 95 481,791 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 documen t 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 algori thm 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 rankin g as developed by IMAC Tea m 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 intern al heat exchangers can vary widely with different evaporators).

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 4 of 99 IMAC S ys tem Test Stand Results Description of Technology An nual COP

Improvement [%] An nual

Energy Improvement [ % ] Cost

Ranking Low=1, High=10 COP/Cost Energy/Cost Comments Improved Conden sers 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 E xternal C ontrolled Variable C om pressor] - 2.4 - 24.1 6.7 - 0.4 - 3.6 Saves significant energy with a small decrease in COP – shows t hat energy is the correct metric. Internal Heat Exchanger 21.4 - 12.3 6.5 3.3 - 1.9 Results depend on evap orator 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 minimal only two points run Improved variable com pressors, som e wit h 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 minimal only two p oints run Superheat Control Algorithm [enabled by use of EXV] 28.6 - 16.5 7.2 4.0 - 2.3 Chang e from 7K to 1K at low loads. Cross charge TXV is similar . Results depend on evaporator design . Best of Best System vers us ARCRPII 15.0 - 36.4

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 5 of 99 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 perform ance. 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, a ll systems used a variable displacement compressor and a manually controlled electric actuated stepper motor expansion valves [EXV]. The system s evaluated here were composed of the following:  C ompressor o IMAC Initial - Externally variable compressor o Alternate Refrigerant CRP Phase II - same o IMAC B - o - B - Improved Efficiency Externally variable compressor  E vaporator o IMAC Initial - Micro - channel design o Alternate Refrigerant CRP Phase II - Enhance d Plate and Fin design o IMAC B - o - B - Improved Efficiency - Micro - channel design  C ondenser- 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 - Im proved Efficiency – Enhanced Micro - channel design

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 6 of 99 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 e nergy is a better metric than COP based on life cycle analysis . We ighting COP gave too high a value to low load conditions where energy consumption is low . Energy consumption weighting be tter assesses the impact on the vehicle . The s teps 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 co mpressor usage/year, calculate kJ of energy consumption for the compressor  We ight 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 bas ed 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 tha t 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 Fin al Report Ap ril, 2007

Page 7 of 99 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 redu ced 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 condit ions 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 in let 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 develop ed 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 cond

itioning system on a test bench  J2766 Standard for Annualized Climate Calculation of System Power Consumption of a mobile air conditioning system Cost vs B enefit 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 conse nsus 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 . Ac knowledgements : 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 Fin al Report Ap ril, 2007

Page 8 of 99 Wewould like to acknowledge the support of t he 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 concern s, their names will not be disclosed here. There were some compon ents 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 W awzyniak Behr Bill Hill GM Dennis Littwin Fujikoki Jan Zeng Denso

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 9 of 99 

## II. System Test Bench Evaluations Technologies Considered : Team 2 began by brainstorming a list of techno logies for consideration. T echnologies were assigned to team members to research. Team meetings , in 2004 , reviewed the potential benefits and ranked the technologies for COP improvement. This proces s resulted in the selection of the most promising techn ologies 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 Fin al Report Ap ril, 2007

Page 10 of 99 1. Improved Heat Exchanger Design s (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 s uppliers were then requested to supply data on improved designs . This data was used by the OEM members to select heat exchanger design s 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 C ompressors 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 . T he individual benefit was ranked based on the GM presentation , b elow is a summary :  ARCRP data suggests a 10 - 20% COP improvement for external controlled compressor as compared to internal control .  Supplier A estimate d the savings to be 15 - 30% depending on loads.  Supplier B estimate d the savings to be 25 -40% depending on loads. The team agreed that 15% wa s 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 Init ially, 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. W ith this assignme nt, 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 sh owing results from no difference to significant advantage for internal control variable . Wh ile 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 agreem ent. 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 Fin al Report Ap ril, 2007

Page 11 of 99 6. Oil Separator This technol ogy is included in some of the optimized compressor efficiency designs and is included in item 1 2 below . It was not considered as a stand alone technology. 7. Ejector System Ejector technology was presented by University of Illinois personnel. W hile th ere 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. Airfl ow 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 t he Best of the Best S ystem 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 1 st 

### generation and a 2 nd 

### generation of improved compressors for evaluation. Some designs include integrated oil separators. 13. Optimized System Control Strategy GM presented ideas for contro lling 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. T he 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 Fin al Report Ap ril, 2007

Page 12 of 99 15. Flash gas removal University of Illinois personnel indicated some potential benefit. Wh ile 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 o n this subject. This technology applies only to electric driven compressors , and for this reason, this item wa s not included for test. 17. Improved evaporator air temperature distribution All testing was done with uniform air temperature inlet to evaporator and con denser. 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. Match ing components to compressor The team decided that since matching system componen ts 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 Compress or 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. Thi s technology was not deemed to provide sufficient improvement to warrant the added complexity.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 13 of 99 Technologies Tested : Technologies s elected 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 variabl e compressors, integrated oil separator Condenser Subcooling Pass Design Subcooling Control Algorithm Superheat Control Algorithm w/EXV Best of Best System The improvement results for indiv idual technologies are compared to a baseline s ystem . Due to testing progression over a two year period, baselines for individual technolog ies are not the same. I n 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 Compresso r  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 (Ph ases I and II), the ACRC (Airconditioning and Refrigeration Research Center) at the University of Illinois at Urbana Champaign was selected for the SAE IMAC test activity. The following gives a brief summary of the test facility. The experimental facil ity, shown below , has two environmental chambers and a

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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, airs ide, and calorimetric chamber balance. Wi th 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 accept ed 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 me ter 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

Dpea Dpen Evap CP M t r Tor

HG Condenser Chamber

Tgo Tgi

B

H CC N Tc n 1

TC Tcai TG

Sp

W Tcro, Pcro Tcao TG M

Dpca Tc n 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 Fin al Report Ap ril, 2007

Page 15 of 99

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 Tcpri Tcpro 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 Fin al Report Ap ril, 2007

Page 16 of 99 Methods of Capacity

and COP Measurement : Evaporator Capac ity 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 h eat – 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 condense r/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 du ri ng 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 n ot a ccurate . As noted above, at the lowest loads only Air Side Capacity is reliable. F or cycling conditions, dry bulb and humidity leaving the evaporator were integrated separately from the high speed data to calculate sensible and latent capacities . Th e issue encountered wa s that the humidity measurement use d 

# air mass flow rate h ai 

# =h (T, RH) h ao 

# =h (T, RH) refrigerant mass flow rate h ri 

### =h (T, P, x) h ro 

### =h (T, P) 

# ± steam glycol 

# air mass flow rate h ai 

# =h (T, RH) h ao 

# =h (T, RH) refrigerant mass flow rate h ri 

### =h (T, P, x) h ro 

### =h (T, P) 

# ± steam glycol

condensate

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 17 of 99 a chilled mirro r which is a slow response devic e itself and the air sample must travel some distance into the device. The chilled mirror is highly accurate at the high relative humidity leaving th e 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 describ ed. 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 1 500W.  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 g raphs 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 (dy namic) 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 conditionError [%] Error: Qe_air compared with Qe_ch Frosting occurred so chamber balance was off

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 18 of 99 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 conditionError [%] Error: Qe_air compared with Qe_ch



### Detail s 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 lo ad points . System capacity was controlled by varying compressor displacement. The evaporator air outlet temperature was controlled to 10oC. T he compressor was run at full capacity at 35°C and 45°C test ambient condition s . Compressor chamber temperature set to 30K above condenser air inlet temperature.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 19 of 99 Controls Discussion : For variable compressor operation: The compressor was controlled with a PW M [Pulse Wi dth Modulated] control. This was used to allow control of the displacement so points with equa l 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 d uring 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 i nput signal for a PID controller to control compressor clutch. The temperature variation band wa s specified as ±1  C. However, this band was allowed to b e widened up to ± 3  C (or other number) to keep the cycle rate lower than 6 cycle s/min ute .

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 20 of 99 Table III - 2 SAE IMAC test matrix (reduce matrix only for component performance comparison test)
Test Name Ambient

temp. [ o C/o F] Compr

speed [rpm] Condenser Evaporator

Inlet air temp. [ o C/ o F] Flow rate

[l/s/CFM] Inlet air

temp. [ o C/ o F] Flow

rate [l/s/CFM ] Simulated

air selection R H

[%] Dew point [ o C] 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 “ 1067/2 260 “ 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 “ 40/8 5 “ “ “ Yellow background for Reduced Matrix

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 21 of 99 The reduced test matrix for fixed compressors included s elected 4000 rpm points . Each fixed compressor had its speed adjusted to match capacity, on average, at I35a, L35a, and M35a. Baseline Variabl e Piston [#4] Fixed Piston [#8] [1.0] Fixed Scroll [#7] [1.65]

900 900 1485 1500 1500 2475 2500 2500 4125 4000 4000 6600 An nualization procedure: For each individual technology, the annualization weighting factors were adjusted based on the points t hat 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 US climate <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 Fin al Report Ap ril, 2007

Page 22 of 99 System S etup 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 tempe rature at the i nlet to the h eat e xchangers. 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 s ystem tested, the a djustment 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 numberTemperature [oC]



### Evaporator inlet air temperatures

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 numberTemperature [oC]



### Condenser inlet air temperatures

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 24 of 99 5 159 313 467 62133

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 Fin al Report Ap ril, 2007

Page 25 of 99 Air side pressure drop: Each time there was a heat exchanger change, t he 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 Fin al Report Ap ril, 2007

Page 26 of 99 Refrigerant charge determination : Charge dete rmination 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 co ndition 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 Am bient

temp. [ o C/o F] Compr

speed [rpm] Con denser Evaporator

Inlet air temp. [ o C/ o F] Flow rate

[l/s/CFM] Inlet air

temp. [ o C/ o F] Flow rate

[l/s/CFM] Simulated

air selection RH

[%] Dew point [ o C] 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. Wh en steady state conditions were achieved data were taken and refrigerant was added in increments of 50g (or 25g). A n 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 1800P [kPa] COP_indoor_avg

DTsub_cro Qindoor_avg Ten Selected charge = 1300 [g] Prcpi Prcpo

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 27 of 99 Oil circulation rate measurement : Before the final performance test, oil circulation rate was measured with two diffe rent 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 me asuring the mass ratio of refrigerant to oil. The measurement results for oil circulation rate and system charge amounts for each system a re given in the following table .

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 28 of 99 Description of Technology Charge Amount [gm] Oil Circulation (%) Comments

90 0 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 Base line 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 Suct ion 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 Subcoo l ing 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 Baseli ne 1125 0.66 1.16 IMAC BoB 1150 1.46 5.95

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 29 of 99 Capacity

and Energy

Test Results : Improved Condensers: Comparison of IMAC heat exchangers to industry average: 

### Condenser Face A rea

0.1 0.2 0.3 0.4Face Area m2

Europe Japan U.S. Group A verage 0,247 m 2avg.all systems AR CRP Phase I Systems

AR CRP Phase II Systems IMAC system 241,600



### The OEM team members developed a condenser component specification and invited condenser manu facturers 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 Fin al Report Ap ril, 2007

Page 30 of 99 S AE IMAC system component table for condenser study Served as baseline Component Feature Cond#3 Cond#6 Cond#7

Condenser Core effective face [mm] W6 26xH386XD16 W5 80xH418xD16 W546xH396x D22 Face area [mm 2 ] 241,636 242,440 216,216 Core volume [mm 3] 3,866,176 3,879,040 4,756,752 Aspect ratio [ -] 0.62 0.72 0.73 No. of tubes 39 (micro - channel) 42 (micro channel) 61 (micro channel) 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] W2 67xH257xD50 Face area [mm 2 ] 68,619 Core volume [mm 3] 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 IR D. Cond#3 was the baseline.

2. Cond#7 was chosen for the final system.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditionQe [kW] Cond#3 baseline Cond#6 Cond#7

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditionCOP [-] 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 conditionWcomp [kW] Cond#3 baseline Cond#6 Cond#7

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditionPcro [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 onl y 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 Comparis on

- 0.4% - 0.8% - 7.4% 6.3% COP comparison:

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 34 of 99 Improved Evaporators: 

### Evaporator Face Area

0.04 0.05 0.06 0.07 0.08Face Area m2

Europe Japan U.S. Group A verage 0,0572 m 2avg. 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 baseli ne 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 Fin al Report Ap ril, 2007

Page 3 5 of 99

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 36 of 99 SA E IMAC system component table for evaporator study Served as baseline Component Feature IM AC Evap#1 IM AC Evap#2 IM AC Evap#3
Condenser (Cond#3) Core effective face [mm] W6 26xH386XD16 Face area [mm 2 ] 241,636 Core volume [mm 3] 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] W2 66xH251xD38 W2 70xH235xD35 W2 67xH257xD50 Face area [mm 2 ] 66,766 63,450 68,619 Core volume [mm 3] 2,537,108 2,220,750 3,430,95 0 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 contro lled 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 Fin al Report Ap ril, 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 conditionQe [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 conditionCOP [-] Evap#1 baseline Evap#2 Evap#3

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditionWcomp [-] Evap#1 baseline Evap#2 Evap#3

Evap#1 temp distribution

0 2 4 6 8 10 12 14 16 18 20

Teao1 Teao11 Teao12

Teao13 Teao14 Teao2

Teao21 Teao22

Teao23 Teao24 Teao3

Teao31 Teao32

Teao33 Teao34 Teao4

Teao41 Teao42

Teao43 Teao44 TC numberTemp [degC] I35a M35a I25a M25a

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 39 of 99 Evap#2 temp distribution

0 2 4 6 8 10 12 14 16 18 20

Teao1 Teao11Teao12

Teao13 Teao14 Teao2

Teao21 Teao22

Teao23 Teao24 Teao3

Teao31 Teao32

Teao33 Teao34 Teao4

Teao41 Teao42

Teao43 Teao44 TC numberTemp [degC] I35a M35a I25a M25a

Evap#3 temp distribution

0 2 4 6 8 10 12 14 16 18 20

Teao1 Teao11 Teao12

Teao13 Teao14 Teao2

Teao21 Teao22

Teao23 Teao24 Teao3

Tea o31 Teao32

Teao33 Teao34 Teao4

Teao41 Teao42

Teao43 Teao44 TC numberTemp [degC] I35a M35a I25a M25a

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 40 of 99 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 Fin al Report Ap ril, 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 commerci al 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:

SA E IMAC system component table for IHX st udy : system No.1 Served as baseline Component Feature without IHX IHX#1 IHX#2
Condenser (Cond#3) Core effective face [mm] W6 26xH386XD16 Face area [mm 2 ] 241,636 Core volume [mm 3] 3,866,176 Aspect ratio [ -] 0.62 No. of tubes 39 (micro - chan nel) No. of Passes 3 (22 - 11 - 6) Mass [kg] 2.1
Evaporator (Evap#1) Core effective face [mm] W2 66xH251xD38 Face area [mm 2 ] 66,766 Core volume [mm 3] 2,537,108 Aspect ratio [ -] 0.94 No. of tubes 39 (micro - channel) No. of Passes two slabs, pass es 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 Fin al Report Ap ril, 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 conditonDP [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 conditonQe [kW] Baseline data 1IHX 2IHX

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditonCOP [-] 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 conditonEta_ihx [-] 1IHX 2IHX

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditonWcomp [kW] Baseline data

1IHX 2IHX



### Only the following points were run: L35a, M35a, I25a, M25a, I15, M15 Weightin g 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 68 3,093 2.77 Energy Comparison

- 10.4% 20.6% - 12.3% 21.4% COP comparison:

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 i s a n enhanced co - tube design . T he second is a compact heat exchanger design. SA E 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 effe ctive face [mm] W546xH396xD22 Face area [mm 2 ] 216,216 Core volume [mm 3] 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] W2 67xH257xD50 F ace area [mm 2 ] 68,619 Core volume [mm 3] 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 i n 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 Fin al Report Ap ril, 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 Fin al Report Ap ril, 2007

Page 47 of 99 SAE IMAC IHX1: Quality/SH sweep @I25a (controlled capacityTen = 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 Fin al Report Ap ril, 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 rea son 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 poi nts 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 conditionCOP / Qe COP Qe

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 49 of 99 Improved Variable Compressors : Initially, two compressor suppliers submitted improved designs that were test ed against the baseline compressor. W hile a n 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 we re reported out in the December, 2005 Interim report. SA E IMAC system component table for compressor study : system No.1 Served as baseline Component Feature Comp#1 Comp#2 Comp#3
Condenser (C ond#5) Core effective face [mm] W6 26xH386XD16 Face area [mm 2 ] 241,636 Core volume [mm 3] 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] W2 6 6xH251xD38 Face area [mm 2 ] 66,766 Core volume [mm 3] 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 con trolled variable displacement External

controlled variable displacement Displacement [cc] ~ 160 ~ 160 ~ 16 0 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 Fin al Report Ap ril, 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 conditionQe [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 conditionCOP [-] Compr #1 Compr #2 Compr #3

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditionWcomp [kW] Compr #1 Compr #2 Compr #3



### Annual results for the 1 st 

### round of improved compressors are reported below with the 2 nd 

### test.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 . SA E 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 effect ive face [mm] W6 26xH386XD16 Face area [mm 2 ] 241,636 Core volume [mm 3] 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] W2 66xH251xD38 Face area [mm 2 ] 66,766 Core volume [mm 3] 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 d isplacement 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 test ed with this system.

2. EXV as expansion device

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 53 of 99 SAE IMAC Compressor comparison: Evaparator CapacityQe

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 conditionQe [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 conditionCOP [-] 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 Fin al Report Ap ril, 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 conditionWcomp [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 Fin al Report Ap ril, 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 M3 5a 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 M2 5a 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 M1 5 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 : E nergy 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 M3 5a 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 M2 5a 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 M1 5 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 Fin al Report Ap ril, 2007

Page 56 of 99 Condenser Sub - cooling Pass Design: At the beginn ing 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 ha d two passes plus sub - cooler except condenser #1 which does not have a sub - cooling pass. Number of tubes in e ach 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 Fin al Report Ap ril, 2007

Page 57 of 99 Three condenser circuiting designs were evaluated.

SA E IMAC system component table for subcooling study Served as baseline Component Feature Cond#1 Cond#3 Cond#5

Condenser Core effective face [m m] W6 26xH386XD16 W6 26xH386XD16 W6 26xH386XD16 Face area [mm 2 ] 241,636 241,636 241,636 Core volume [mm 3] 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 P asses 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] W2 66xH251xD38 Face area [mm 2 ] 66,766 Core volume [mm 3] 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 total surface [%] First pass Second pass Sub - cooling pass Condenser #1 26 13 0 0 Condenser #3 22 11 6 18 Condenser #5 18 9 12 44

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 Fin al Report Ap ril, 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

I4 5 L45 I35a L35a M35a I2 5a L25a M25a I1 5 L15 M15 Test conditionQ [kW] Q - Cond#1 Q - Cond#3 Q - Cond#5

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditionCOP [-] 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 conditionWcomp [kW] Wc om p - Cond#1 Wc om p - Cond#3 Wc om p - Cond#5

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 ad vantage.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 loopQe [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. T eam 2 feels that there still may be merit but a b etter method of separating the gas from the liquid in the receiver will be necessary to better evaluate this technology. Since there was not s ignificant improvement noted at these two screening points an annual number was not calculated.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 63 of 99 Superheat Con trol 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 stud y one:

SA E IMAC system component table for superheat study Component Feature
Condenser (Cond#5) Core effective face [mm] W6 26xH386XD16 Face area [mm 2 ] 241,636 Core volume [mm 3] 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] W2 66xH251xD38 Face area [mm 2 ] 66,766 Core volume [mm 3] 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 Fin al Report Ap ril, 2007

Page 64 of 99 The superheat was then varied from 2 to 10 K . 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 20EVAP 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 Fin al Report Ap ril, 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 5 22,346 2.96 Energy Comparison

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

### Superheat control study two:

SA E 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 [mm 2 ] 216,216 Core volume [mm 3] 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] W2 67xH257xD50 Face area [mm 2 ] 68,619 Core volume [mm 3] 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] ~ 16 0 Mass (including pu lley) [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 Fin al Report Ap ril, 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 Fin al Report Ap ril, 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 Fin al Report Ap ril, 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  Sup erheat Control Algorithm w/ EXV The B - o - B System was run to match ARCRPII evaporator air off temp eratures and also at elevated temperatures.

SA E 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 [mm 2 ] 216,216 Core volume [mm 3] 4,756,7 52 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] W2 67xH257xD50 Face area [mm 2 ] 68,619 Core volume [mm 3] 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] ~ 16 0 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 Fin al Report Ap ril, 2007

Page 69 of 99 SAE IMAC compared to Phase II: Evaparator CapacityQe Note: Maximum Qe for 45/35degC and controlled Qe for 25/15degC

0 1 2 3 4 5 6 7 8I60aI60bI60cI45L45M45H45I50aI35aL35aM35aH35aI35bL35bM35bH35bI40aI25aL25aM25aH25aI40bI25bL25bM25bH25bI25cL25cM25cH25cI30I15L15M15H15 Test conditionQe [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 CapacityQe Note: Maximum Qe for 45/35degC and controlled Qe for 25/15degC

0 1 2 3 4 5 6I60aI60bI60cI45L45M45H45I50aI35aL35aM35aH35aI35bL35bM35bH35bI40aI25aL25aM25aH25aI40bI25bL25bM25bH25bI25cL25cM25cH25cI30I15L15M15H15 Test conditionCOP [-] Phase II A RCRP IMAC B-O-B match phase II IMAC B-O-B 10 deg C

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 6I60aI60cL45H45I35aM35aI35bM35bI40aL25aH25aI25bM25bI25cM25cI30L15H15 Test conditionWcomp [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. AR CRP Base IMAC B -o - B Teao= A RCRP 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 Fin al Report Ap ril, 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.

IM AC Base Co mpr energy IM AC B - o - B

Compr energy IM AC Base COP IM AC B - o - B

COP 663, 4 95 481,791 2.20 3.39

27.4% 54.3%

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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.

SA E 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 [mm 2 ] 216,216 Core volume [mm 3] 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] W2 67xH257xD50 Face area [mm 2 ] 68,619 Core volume [mm 3] 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 ~16 0 Mass (i ncluding 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 Fin al Report Ap ril, 2007

Page 73 of 99 SAE IMAC Variable vs. Fixed compressor: Evaparator Capacity Qe

0 1 2 3 4 5 6 7 8I45L45H45I35aL35aM35aH35aI25aL25aM25aH25aI15L15M15H15 Test conditionQe [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 conditionCOP [-] IMAC TXV data (used as fixed compressor baseline) IMAC comp#8 (fixed) data IMAC comp#7 (fixed) data

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 conditionWcomp [kW] IMAC TXV data (used as fixed compressor

baseline) IMAC comp#8 (fixed) data IMAC comp#7 (fixed) data



### An nualization: IM AC EXV IM AC

TXV Fixed

Compr #7 Fixed

Compr #8 IM AC

EXV IM AC

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% bas e 20% 27% - 3% base - 11% - 19%

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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: T he 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 (va rying 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 - Pr od Honeywell

Jetta I-MAC / Load

reduction I-MAC /

Optimized Ref. I-MAC /

Production.Degrees F.

City Traffic Road

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 minComfort 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 minComfort Rating



### Driver Rt. Front Lft. Rear Rt. Rear Comfort

Team 7, 10 and 12 rides System problems for 8 and 11 rides Al l 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 Fin al Report Ap ril, 2007

Page 77 of 99 S hown 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 , s howing equal performance. Ave rage 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 Fin al Report Ap ril, 2007

Page 78 of 99 Th e vehicle development plan is shown below which progressed with 3 generations of algorithm improvement, with a final verification of the 3 rd 

### / 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 co mpressor and expansion valve to optimized COP control. All other co mponents were original equipment parts from the 2005 Cadillac STS. The first generation of s oftware 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 regio n the compressor displacement was controlled to the maximum displacement, the expansion valve was controlled to maintain optimal superheat for maximum capacity , a nd the air recirculation door was set to 0% outs ide 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 compresso r 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 ou tside air for occupant comfort , and the EXV is controlled to maintai n an optimal superheat for maxi mum COP. Task Responsibility Completion Comments Obtain Vehicle Bill Hill Complete - see Note 1 GM will provide 2 Cadillac STS Vehicles

(Ef ficiency 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 S ys 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 Da ve 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 - Gold en, CO Bill Hill Complete Soak & cool down test for CAE Model Corr Vehicle Algorithm dev Team 2 Complete In Detroit Wi nd Tunnel Test Team 2 Sept 20 - 21, 2005 Scrubbed due to instrumentation problems Develop Future Vehicle Test Plans Team 2 Com plete 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 Fin al Report Ap ril, 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 MicroAu tobox 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 (H WI O) Layer which reads the data inputs and provides the control ler outputs, the Software Integration (SWI L) 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 Fin al Report Ap ril, 2007

Page 80 of 99 The software model include s 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. T he compressor power was calculated in generation one o riginally using a normalized load calculation and the refrigerant flow rate , which was calculated by multiplying compressor displacement (Disp), duty cycle, speed (V c 

### ), suction density (ρ s 

### ) and volumetric efficiency (η v 

### ) 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. T he d ata acquisition system (DAS) was als o 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 ele ctronic 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 th e 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 ‘E DS method corrected’.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 L25bCapacity ( 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 diffi cult to separate and estimate with reasonable accuracy.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 82 of 99 The following plot shows the effect of changes on compressor speed and evaporator air exit temperature (EAT) on COP. Wh ile 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 4500Compressor Speed (rpm)

COP EAT RPM 

### Generation two: Following the Texas ride, the vehicle was returned to Detro it to switch compressors to the compressor that performed best at the University of Illinois ACRC la bs. At this time the method (2 nd 

### Algorithm version) for estimating mass flo w rate and compressor power was ch anged to use linear regression. The regression was based off of the air flow rate on the condenser (AFR c 

### ), the air temperature entering the condense r (T cai 

### ), the air temperature exiting the evaporator measured at the nozzle (T en 

### ), the compressor speed (V c 

### ), the discharge pressure (P cpro 

### ), the compressor displacement control current (CompCurr) and the air flow rate over the evaporator (AFR e 

### ).

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 83 of 99 The re gression 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 (K p 

### ) using the equation below.

This model was found to matc h 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 FlowRate Estimation

R2= 0.9542

0 10 20 30 40 50 60

0 10 20 30 40 50 60 IMAC Bench DataCalculated Value 7 6 5 4 3 2 1 0 

## a

AFR a CompCurr a P a V a T a T a AFR a M

e cpro c en cai c r 

##               

7 6 5 4 3 2 1 0 

# b

AFR b CompCurr b P b V b T b T b AFR b P

e cpro c en cai c comp 

#               

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 DataCalculated Value



### The vehicle was taken to Phoenix in the summer , 2006 and high temperature testing was performed. Int erestingly, 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 indica ted 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 el iminate dynamics that may have a ffected 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 3COP ( - ), Capacity (kW)

Superheat Setpoint Actual Superheat Evaporator Capacity COP Evaporator Capacity COP

Actual Superheat Superheat Setpoint

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 85 of 99 Generatio n three: After the Phoenix Alternative Refrigerant System Symposium, the Algorithm was then modified (3 rd 

### Version) , to incorporate a physically based approach. In this approach the original mass flow rate equation which calculated mass flow rate by multi plying 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 ( P d 

### /P s 

### ) , compressor pressure difference *(P d 

### - P s 

### ) , compressor speed (V c 

### ) , 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 C 1 

### through C 7 

### . Kv 

### = (1 - C 0 

### *P d 

### /P s 

### )*(C 1 

### *(P d 

### - P s 

### )+C 2 

### *CompCurr+C 3 

### *V c 

### )+C 4 

### +C 5 

### *P d 

### +C 6 

### *P s 

### +C 7 

### *CompCurr A regression for compressor power was done in the same fashion calculating a different term (K p 

### ) using the equation below. K p 

### = (1 - D 0 

### *P d 

### /P s 

### )*(D 1 

### *(P d 

### - P s 

### )+D 2 

### *CompCurr+D 3 

### *V c 

### )+D 4 

### +D 5 

### *P d 

### +D 6 

### *P s 

### +D 7 

### *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. I t should be noted that the regression predicts high values of COP. Regression Results for COP vs. IMA C Bench Data

0 1 2 3 4 5 6 7 8 9 10

I60a I60b I60c I45 L45 M45 I50a I35a L35a M35a I35b L35b M35b I40a I25a L25a M25a I40b I25b L25b M25b I25c L25c M25cCOP ( - )

COP - bench COP - vehicle regression 

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

M25a, L25b, M25b, L25c, and M2 5c.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 86 of 99 Regression Results for Capacity vs. IMAC Bench Data

0 1 2 3 4 5 6 7 8

I60a I60b I60c I45 L45 M4 5 I50a I35a L35a M35a I35b L35b M35b I40a I25a L25a M25a I40b I25b L25b M25b I25c L25c M25cCapacity (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

I60a I60b I60c I45 L45 M45 I50a I35a L35a M35a I35b L35b M35b I40a I25a L25a M25a I40b I25b L25b M25b I25c L25c M25cCompressor 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 mat ched 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 Fin al Report Ap ril, 2007

Page 87 of 99 displacement was h igher 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 L25bCOP ( - )

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 L25bCapacity ( kW )

Capacity bench data Capacity car regression

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 L25bWork (kW)

Wcomp bench data Wcomp car estimated

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 2 nd 

### Algorithm version gave conflicting results from Bench testing. The study also examined the effects of re - circulated air control and e vaporator temp erature control on compressor work. The test poi nts in the vehicle were at IMAC conditions I25b, L25b, and M25b as wel l 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 superh eat 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 Ti me (min)Superheat (°C) Measured Superheat Superheat Setpoint

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 Fin al Report Ap ril, 2007

Page 91 of 99 Effect of superheat on Evaporator Capacity at constant Compressor PW M 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 conditio n, 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 Fin al Report Ap ril, 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 cond itions 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 ma intain 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 Fin al Report Ap ril, 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 (2.4k W) , 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 Fin al Report Ap ril, 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 ou t 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 displace ments 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 A ir 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 Fin al Report Ap ril, 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 Ai r Recirculation Percentage (%)Compressor Power (kW) I25 L25 M25 Phoenix

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 2007

Page 96 of 99 Verification of capacity regression [Measured vs Algorithm]: Adding the refrigerant flow meter allowed for an accurate measure ment 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 fi nal 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 re gression 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 47Capacity (kW)

Q regression Q measured

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 iterati on was a physically based approach implemented by EDS, the second was a linear regression approa ch implemented , and the thi rd 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 L25bCOP ( - )

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 impl emented 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 cal culated using the bench data to account for compressor efficiency.

### IMAC Team 2 - Improved Efficiency Fin al Report Ap ril, 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 L25bCapacity ( kW )

Vehicle EDS equations Vehicle original regression Vehicle final regression Capacity bench data Vehicle air side dry 

### Ai rside 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 b elow due to the lack of the latent load component. To calculate the capacities the air flow rate (A FR c 

### ) 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 Fin al Report Ap ril, 2007

Page 9 9 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 L25bPower (kW)

Wcomp vehicle EDS equations Wcomp vehicle original regression Wcomp vehicle final regression Wcomp bench data 

### Conclusion: The fi nal 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.