FINAL REPORT 

Demonstration of an Energy-Efficient 

Secondary Loop 

HFC-152a Mobile Air Conditioning System 

By 

 

Mahmoud Ghodbane 

Timothy D. Craig 

James A. Baker 

Delphi Thermal Systems Division 

Delphi Corporation 

for 

U.S. Environmental Protection Agency 

c/o Kristen Taddonio, Project Officer 

Strategic Climate Projects 

 Climate Protection Partnership Division 

1200 Pennsylvania Avenue, MC 6202 J 

Washington, D.C. 20460

Order No. EP07H001055 

 

 December 17, 2007 
2

Page 

Table of Contents 2 

 

List of Figures 4 

List of Tables 5

 

I. INTRODUCTION 6

 

II. BACKGROUND 7 

 

III. EPA PROJECT 8 

 3.1 Project Objective 8 

3.2 Testing and Demonstration 8 

3.3 Timing 9 

3.4 Design Guidelines 10 

IV. PRODUCTION R-134A VEHICLE AND SYSTEM 

 DESCRIPTION 10 

V. R-152A SECONDARY LOOP SYSTEM DESCRIPTION 11 

VI. SECONDARY LOOP SYSTEM COMPONENTS 13 

 6.1 Compressor 13 

 6.2 Condenser 14 

 6.3 Chiller 14 

 6.4 Refrigerant Controls 16 

 6.5 Cooler 17 

 6.6 Pump 18 

 6.7 Coolant 18 

VII. PERFORMANCE COMPARISON 21 

7.1 Test Description and Methodology 21

7.2 Performance under High Ambient Conditions 22 

7.3 Road Performance under High Loads 24 

7.4 Performance under Medium to Mild Ambient Conditions 26 
3

 VIII. ENERGY COMPARISON 32 

8.1 At Full Capacity 32 

8.2 At Controlled Capacity 33 

8.3 Fuel Consumption Analysis 37 

XI. SECONDARY LOOP DESIGN GUIDELINES 41 

9.1 Overall System 41 

9.2 Potential Design Improvements 42 

9.2.1 System Mass 43 

9.2.2 Heat Exchanger Effectiveness 44 

9.2.3 Component Package Size 46 

9.2.4 Compressor Efficiency 46 

9.2.5 Control Strategy 46 

9.2.6 “Cold” Storage Systems 47 

X. SUMMARY AND CONCLUSIONS 47 

 

ACKNOWLEDGEMENT 48 

REFERENCES 48 
4

List of Figures 

 Page 

Figure 1 2006 Opel Astra 10 

Figure 2 Opel Astra R-134a A/C System 11 

Figure 3 Solid Model of Secondary Loop System 12 

Figure 4 Opel Astra R-152a Secondary Loop System 

 Lay Out 12 

Figure 5 Typical Chiller 14 

Figure 6 Chiller Options 15 

Figure 7 Chiller Heat Transfer Performance 15 

Figure 8 Chiller Coolant Pressure Drop 16 

Figure 9 Effect of Refrigerant Chiller Out Super Heat on 

 Chiller Effectiveness 17 

Figure 10 Typical Cooler 17 

Figure 11 Cooler Heat Transfer Performance 18 

Figure 12 Coolant Pump 19 

Figure 13 Secondary Fluids (Coolants) 20 

Figure 14 Performance Comparison at High Loads 22 

Figure 15 Performance Comparison at High Loads 23 

Figure 16 Road Performance at High Loads 24

Figure 17 Road Performance at High Loads 25

Figure 18 Subjective Comfort Evaluation 26 

Figure 19a Secondary Loop Capacity Control in Outside 

 Air Mode (Overcooling Reduction) 27 

Figure 19b Secondary Loop Capacity Control in Outside 

 Air Mode (Overcooling Reduction) 27 

Figure 20a Secondary Loop Capacity Control in Outside Air 

 Mode, Medium Loads (Overcooling Reduction) 28

Figure 20b Secondary Loop Capacity Control in Outside Air 

 Mode, Medium Loads (Overcooling Reduction) 29

Figure 21a Secondary Loop Capacity Control in Outside Air 

 Mode, Low Loads (Overcooling Reduction) 30 

Figure 21b Secondary Loop Capacity Control in Outside Air 

 Mode, Low Loads (Overcooling Reduction) 31 

Figure 22 Coefficient of Performance Comparison at Design 

 Conditions 32 

Figure 23 Coefficient of Performance Comparison in Outside 

 Air Mode 33 

Figure 24a Coefficient of Performance Comparison at Medium 

 Loads 34 

Figure 24b Coefficient of Performance Comparison at Low 

 Loads 34 

Figure 25a Compressor Power vs Capacity at Road Load 35 

Figure 25b Compressor Power vs Capacity at Idle 36 
5

Figure 26a Speed Correction Schedule (SC03) 37 

Figure 26b New European Driving Cycle (NEDC) 37 

Figure 26c Highway Fuel Economy Test (HWET) 38 

Figure 27 Sample Inputs to the “Advisor” Software 38 

Figure 28 Vehicle and A/C Data 39 

Figure 29 Overall Vehicle Fuel Consumption with A/C 40 

Figure 30 A/C Fuel Consumption 41 

Figure 31 High Ambient Warm Up with Cold Storage 44 

Figure 32 Compressor Power vs Effectiveness 45 

List of Tables 

Table 1 Test Matrix 21 

Table 2 Air Flow Rates 29 

Table 3 System Mass 43 

 
6

I. INTODUCTION 

All currently manufactured automobile and light duty truck vehicle air conditioning 

systems use HFC-134a as the refrigerant. HFC-134a replaced CFC-12 in the early 

1990’s due to the negative impact of CFC’s on the Earth’s ozone layer. While other 

refrigerants were studied at that time as possible replacements, HFC-134a was 

chosen for its combination of refrigerating properties and safety characteristics, 

despite indications that its global warming potential (GWP) and resulting climate 

impact could one day become a concern. That concern was realized with the recent 

passage of legislation in the European Community that requires the use of 

refrigerants with global warming potentials less than 150 in all new-type vehicles 

starting in 2011 and in all new vehicles by 2017. 

The chemical industry has been working to synthesize new chemicals and mixtures 

of chemicals that might one day prove useful, but has met with limited success to 

date. Only two commercially available refrigerants exist today that have a GWP 

below the required threshold, these being carbon dioxide (R-744) and HFC-152a (R152a). Recently, DuPont and Honeywell have proposed HFO1234yf as R-134a 

replacement. The automotive industry will likely have to choose between these 

alternatives very soon to meet the EU timeline for implementation. 

R-744 is a refrigerant that requires completely new components, engineered to 

withstand its inherently very high system pressures, which results in high capital 

expenditures and unique system development challenges, especially to limit 

refrigerant leakage. In addition, safety concerns associated with the health effects of 

excessive leakage into the passenger compartment will likely require engineering 

mitigation, at least in the United States [1]. 

R-152a is very similar to the current R-134a in terms of pressure, refrigerating 

characteristics, and compatibility with currently used system components and 

materials, making it easier and less expensive to implement. It does, however, 

suffer from the disadvantage of being mildly flammable, which can be successfully 

mitigated by applying the refrigerant to a secondary loop system, in which the 

refrigerant is contained completely within the engine compartment where it chills a 

secondary fluid (similar to engine coolant used for heating) that is then pumped into 

the passenger compartment to cool cabin air [1]. This requires additional 

components, but is still a lower cost and lower commercial risk option than R-744. 

As far as HFO1234yf is concerned, the jury is still out on its candidacy as R-134a 

replacement. 

The US Environmental Protection Agency has recognized the environmental 

opportunities R-152a can offer to the vehicle air conditioning industry and the 

importance of demonstrating this technology in a vehicle system that highlights 

energy efficiency and commercially viable components. To these ends, this 

demonstration program was created. 
7

II. BACKGROUND 

Delphi has a long-standing history of secondary loop systems development, 

beginning with the writing of an internal document on the use of flammable 

refrigerants in 1997 [2]. This led to constructing a secondary loop bench test system 

in 1998 for hands-on system development. Following successful laboratory testing, 

vehicle systems were built, tested and demonstrated publicly. The idea was to 

highlight the potential for using a secondary loop system to enable the use of a 

mildly toxic or mildly flammable refrigerant as a means of satisfying potential 

legislation requiring the use of low GWP refrigerants [3]. 

The first demonstration was a proof-of-concept exercise using identical 1999 

Chevrolet Malibu’s, one equipped with a secondary loop R-134a system and the 

other a regular production, direct expansion R-134a system. These vehicles were 

demonstrated at the 1999 Phoenix Forum (currently the SAE Alternate Refrigerant 

Systems Symposium) during which participants were invited to ride in both vehicles 

and see if they could identify the secondary loop vehicle. Although it had been 

postulated that cooling rate of the secondary loop system would lag behind that of 

the production system due to the intermediate coolant, the participant ratings 

showed equivalent cooling performance and comfort. Whereas a temperature lag is 

measurable during cool down, it is not perceptible by the vehicle occupants. This is 

due to the rapid temperature change associated with the cooling down of a hot 

vehicle and the fact that the bulk of the coolant is stored under-hood where its 

temperature remains near ambient, unlike the passenger cabin, which receives 

significant additional heating from solar radiation. This demonstration followed the 

publication of an SAE paper on the use of flammable refrigerants [4]. 

At the 2003 SAE Alternate Refrigerant System Symposium, Delphi and Volvo jointly 

demonstrated a Volvo XC90 (SUV) R-152a secondary loop system with proof-ofconcept (pre-prototype) components [5].ref. This testing was quite successful with 

the cooling performance of the XC90 being among the best of the vehicles tested. 

At the 2007 SAE Alternate Refrigerant System Symposium, Delphi demonstrated 

several potential alternatives in direct expansion systems being proposed by the 

chemical companies. As part of the current project with the EPA, Delphi also 

successfully demonstrated an R-152a secondary loop system, built and tested by 

Delphi with funding provided by the US EPA [6]. All components were commercially 

viable in a system designed to deliver energy efficient operation. 

In addition to the activities noted above, Delphi engineers have published numerous 

SAE technical papers on the topics of alternative refrigerants and A/C system 

energy efficiency [7]. These are cited in the reference section of this report. 

Secondary Loop R-152a Demonstration Vehicle Project Overview (annotated 

original submission to EPA) 
8

III. EPA PROJECT 

The Project Plan, broadly described, consists of measuring the cooling and energy 

performance of an A/C system on a donated 2007 Opel Astra. This is followed by 

retrofitting that same vehicle with a commercially viable secondary loop HFC-152a 

system. The vehicle was retested for cooling performance and energy, and then 

demonstrated the vehicle with the secondary loop system at the 2007 SAE Alternate 

Refrigerant Systems Symposium at their mid-July meeting in Scottsdale, Arizona. 

3.1 Project Objective 

System design consists of fully understanding the current physical and operational 

characteristics of the Astra and applying that knowledge to designing the necessary 

new components to create the desired secondary loop system. Design details and 

rationale will be provided to a Peer Review Team for input and concurrence. At this 

time, the Peer Review Team consists of William Hill (GM), Hans Fernqvist (Volvo), 

Stephen Andersen (EPA), and Ward Atkinson (SAE). Additional globally recognized 

experts (e.g., R. Monforte (Fiat) and J. Rugh (NREL) will be invited to join. A Peer 

review meeting (in person or via teleconference) will be held to discuss initial 

component and system designs, as well as the intended test methodologies to 

quantitatively measure system characteristics. A second meeting(s) will be held 

during the testing phase for those interested in viewing the data of the secondary 

loop system. Outputs from these meetings that would contribute to improved system 

performance will be incorporated into the design guide. 

Attached in appendix A: Power point file (EPA_R152a_Project_overview.ppt) 

containing diagrams that both reference, and further describe, the new components 

to be designed, developed and incorporated into the system. In appendix B: The 

first Peer Review Team Meeting Presentation. The link to the presentation given at 

2007 SAE Alternate Refrigerant Symposium given in Scottsdale, AZ is below: 

http://www.sae.org/events/aars/presentations/2007aarsvehicles.pdf

3.2 Testing and Demonstration 

Also included in the attached PowerPoint file is the summary of the testing protocol 

for the two system architectures. These are, of course, identical except for the 

necessary determination of proper system charge for the secondary loop system 

prior to testing. 
9

3.3 Timing 

Thermal R&D Project Plan

Project Title: Secondary Loop R-152a Demonstration Vehicle Project Number(s):

Project Leader: Mahmoud Ghodbane (CL) (CL) (CL)

Technical Team: James Baker

Shawn Caple Approval Date:

Donald Enzinna

William Kumpf Revision Date:

Xiaoxia Mu

David Polisoto

Paul Droman

Timothy Craig

Lindsey Leitzel

Program Event Start Finish Status Comments

Planned 

(mm/dd/yy)

Actual 

(mm/dd/yy) Planned Actual

1 Preliminary System Design and Testing Plan 02/01/07 02/01/07

2 Sub-System Modeling and Optimization 02/15/07 02/15/07 Extended Optimization, No IHX

3 Bench Test of Preliminary Design Secondary Loop Sub-System (SLSS) 03/12/07 03/12/07 Matched to Optimization; Pump & Heat Ex.

4 Vehicle Packaging of SLSS and Components 02/26/07 02/26/07 CAD study supported by vehicle build

5 Peer Design Review #1 03/19/07 04/16/07 At SAE Congress; component data avaiable

6 Select Final SL Design 03/20/07 03/20/07

7 Peer Design Review # 2 -- Precluded based on accelerated timing

8 Demonstration Vehicle -- Tunnel Testing

8a Ship Vehicle to Lockport, NY Technical Center (LTC) 03/19/07 03/19/07 Start date for Vehicle work

8b Baseline Test Vehicle with R-134a System 04/09/07 04/23/07

8c Update Vehicle with SLSS 04/16/07 05/14/07

8d Test Vehicle with SLSS 04/23/07 06/01/07

8e Update SLSS Based on 8d tests 05/07/07 06/12/07 Instrument for road testing

8f Retest Vehicle with SLSS Updates 05/16/07 -- Not necessary

8g Validate Vehicle Performance In Tunnel 05/16/07 06/18/07 Controls calibration

9 Peer Performance Review 05/24/07 04/12/07 Multiple reviews with GM on data.

10 Demonstration Vehicle -- Road Testing

10a Prepare Vehicle for Road Test and Data Collection 05/28/07 06/12/07

10b Ship Vehicle to Phoenix, AZ 06/04/07 06/22/07

10c Vehicle Road Tests 06/19/07 07/17/07 SAE ARSS Evaluations

11 Design Manual

9a Initial Draft 08/21/07 07/02/07 Review and summary at SAE ARSS

9b Interim Draft 09/15/07 --

9c Final Manual 09/28/07 10/29/07 Adjusted target based on vehicle usage

3/1/2007

EP07H001055, EPA PO#

07/12/07

03/19/07 04/16/07

04/20/07 04/20/07

--

03/16/07 04/20/07

03/09/07 05/11/07

02/21/07 02/26/07

03/09/07 04/20/07

03/30/07 04/23/07

04/13/07 05/04/07

05/18/07

--

06/22/07

07/13/07

05/30/07 06/15/07

--

11/28/07

06/22/07 07/19/07

08/21/07 07/17/07

09/15/07

09/28/07

11/9/2007

06/08/07

05/11/07 06/15/07

05/24/07

04/20/07 05/31/07

04/27/07 06/08/07

05/18/07

© 2007 Delphi Corporation, Inc. 
10

3.4 Design Guidelines 

As noted above, design guidelines to maximize secondary loop R-152a system 

energy efficiency will be collated into design guidelines to accompany the final 

report. 

As requested in the statement of work, EPA will have all property rights and 

ownership of Design Manual. It is understood that property rights and ownership of 

the Design Manual means the manual itself, as a tangible publication, and not 

ownership of the intangible and potentially patentable ideas disclosed therein. 

IV. PRODUCTION R-134a VEHICLE AND SYSTEM DESCRIPTION 

A 2006 Opel Astra production vehicle was used in the investigation. The actual 

vehicle is shown in figure 1. It is equipped with a manually controlled air conditioning 

system. 

 

The production R-134a system is laid out schematically in figure 2. The system is 

comprised of an R-134a refrigerant circuit consisting of components such as 

compressor, condenser, expansion device, and refrigerant lines located under-hood 

and an evaporator positioned inside the HVAC case, which is situated in the 

passenger compartment of the vehicle. 

Figure 1 2006 Opel Astra (1.3 liter gasoline engine) 
11

V. R-152a SECONDARY LOOP SYSTEM DESCRIPTION 

It was shown [2] that R-152a (1,1-Difluoroethane) is an excellent refrigerant. 

Obviously, the only negative characteristic of R-152a is its mild flammability. One 

way to overcome the safety aspect of the mildly flammable refrigerant is to use it in 

conjunction with a secondary loop cooling system. Figure 3 depicts a solid model of 

a Secondary loop system. 

 

The conventional evaporator is replaced by a secondary fluid (coolant) to air heat 

exchanger called a cooler. The secondary loop system uses an R-152a refrigerant 

direct expansion cooling system (primary system) to cool the coolant via a 

refrigerant to coolant heat exchanger called a chiller. The cold coolant is then 

pumped trough the cooler inside the passenger compartment of the vehicle to 

provide the necessary cooling. 

In order for a Secondary loop system to be viable alternative, the heat transfer 

resistance resulting from the added level of heat exchange must be minimized 

through proper selection of the coolant and system design. In addition, the 

packaging and mass of the pump, reservoir, and chiller must be optimized. Unlike 

the primary refrigerant, it is necessary for the secondary fluid to be non 

Figure 2 Opel Astra R-134a A/C System Layout 

Suction Line

Condenser

Evaporator & TXV

Liquid Line

Compressor

Front of Dash 

© 2007 Delphi Corporation, Inc. 
12

 

 

 

 Figure 4 Opel Astra R-152a Secondary Loop System Layout 

Chiller & 

TXV

Cooler

Pump

Compressor

Condenser

Reservoir

Suction Line

Liquid Line

Chiller & 

TXV

Cooler

Pump

Compressor

Condenser

Reservoir

Suction Line

Liquid Line

Front of Dash 

Figure 3 Solid Model of Secondary loop system 

© 2007 Delphi Corporation, Inc. 

© 2007 Delphi Corporation, Inc. 
13

flammable and non-toxic because it is circulated inside the passenger compartment 

of the vehicle. There are a variety of commercially developed single-phase heat 

transfer fluids, which could be used as coolants. In this investigation, ethylene glycol 

/ water solution was chosen because it is common, wildly used, and validated for 

automotive applications. The selection of the percentage of the ethylene glycol used 

in the coolant is detailed later in the report. 

Figure 4 shows in a schematic fashion the lay out of the Opel Astra Secondary loop 

system. The system encompasses a primary refrigerant circuit consisting of 

compressor, condenser, and an expansion device coupled with the coolant circuit 

through the chiller. In this configuration, the chiller is located up front, near the 

compressor and cools the circulating coolant. The amount of coolant flow to the 

cooler inside the vehicle is regulated by a variable speed pump. When comparing 

the baseline lay out shown in figure 2 and the secondary loop system lay out in 

figure 4, the primary refrigerant suction and liquid lines of the secondary loop system 

are much shorter than the baseline system. The short suction line introduces low 

refrigerant pressure drop, which translates into performance improvement. 

VI. SECONDARY LOOP SYSTEM COMPONENTS 

As shown previously in figure 3 the secondary loop system is comprised of primary 

refrigerant circuit consisting of compressor, condenser, and expansion device 

coupled with a secondary fluid circuit via a chiller. The secondary fluid circuit is 

composed of a reservoir, a pump, and a cooler taking place of the evaporator inside 

the HVAC module. The chiller, which is thermally equivalent to an evaporator, is the 

link between the refrigerant circuit and the secondary fluid circuit. 

6.1 Compressor 

As in the direct system, the compressor still is the heart of the secondary loop 

system, and consumes most of the energy. The sizing and the selection of the 

compressor is critical to any system and in the case of secondary loop system, it is 

more significant because of the added thermal resistance introduced by the 

secondary fluid. 

If equivalent cooling capacity and similar compressor technology is used for the 

baseline and the secondary loop system, the displacement of the secondary loop 

compressor is expected to be higher (10 to 15%) than the baseline compressor; 

however, this increase can be reduced to a single digit if design and efficiency 

improvements of the compressor are realized. A theoretical evaluation of the 

displacement increase can be expressed by the following equation: 

















































∆

∆

















=

















−

−

−

−

−

− −

R a

R a

vol R a

vol R a

evap R a

evap R a

suct R a

suct R a

R a

SL R a

RPM

RPM

h

h

Disp

Disp

152

134

152

134

152

134

152

134

134

152

η

η

ρ

ρ
14

Where Disp is the compressor displacement. ρ , ∆h, η, and RPM represent refrigerant 

density, enthalpy difference across the compressor, isentropic efficiency of the 

compressor, and compressor speed respectively. Using the saturation properties of 

the refrigerants at 41(5), 50(10), 59 15), and 81 0

F (27 0C) compressor in 

temperatures with the assumption that both compressors have identical speed and 

isentropic efficiency, on the average the R-152a compressor displacement is higher 

by 11.5%. 

 

6.2 Condenser 

Identical condenser and fan were used for the baseline and the secondary loop 

system. As in any mobile air conditioning, front air management is important for 

optimum system performance especially during engine idling situations. When the 

vehicle is idling, re-circulation of the warm air raises the average air temperature 

entering the condenser and therefore reduces the capacity of the condenser. The recirculation of warm air may be also caused by insufficient clearance between the 

engine and the front-end module in which case the warm air from the front-end 

module is partially reflected back to the condenser. An adequate shrouding of 

condenser, fan, and radiator module was employed to limit warm air re-circulation. 

Sufficient shrouding of the front-end module can prevent re-circulation of the warm 

air discharged from the module. 

6.3 Chiller 

 

Figure 5 Typical Chiller 

Expansion Device 

© 2007 Delphi Corporation, Inc. 
15

The chiller is a refrigerant to secondary fluid (coolant) heat exchanger. In a 

secondary loop system, the refrigerant evaporation occurs in the chiller instead of 

the evaporator for a baseline system. Technically the chiller is an evaporator using 

the coolant as heat transfer medium instead of the air. The chiller used in this 

investigation is shown in figure 5. 

0

100

200

300

400

500

600

0 1 2 3 4 5 6

Coolant Flow (gpm)

Heat Transfer (Btu/min.)

3.78 7.57 11.36 15.14 18.93

(lpm)

8.78

7.03

5.27

3.51

1.75

10.54

(kW)

T coolant in = 50 0F (10 0C)

Currently at the prototype stage, the chiller is a brazed aluminum plate type with 

enhanced heat transfer surfaces for optimum heat transfer and pressure drops. 

Different shapes and designs are portrayed in figure 6. 

Figure 6 Chiller Options 

 Figure 7 Chiller Heat Transfer Performance 

© 2007 Delphi Corporation, Inc. 

© 2007 Delphi Corporation, Inc. 
16

0

1

2

3

4

5

6

0 1 2 3 4 5 6

Coolant Flow (gpm)

(psi)

Refrigerant

Coolant

3.78 7.57 11.36 15.14 18.93

(lpm)

34.47

27.58

20.68

13.80

6.89

41.37

(kPa)

T coolant in = 50 0F (10 0C)

Figures 7 and 8 depict the thermal performance of the chiller used in this 

investigation. Figure 7 includes the heat transfer at fixed coolant in temperature and 

different coolant flow rates. Figure 8 shows the variation of the refrigerant and the 

coolant pressure drops with the coolant flow rate at fixed chiller coolant in 

temperature. As expected, the heat transfer increases with coolant flow rate. The 

data has also shown that the heat transfer drops off when the chiller coolant in 

temperature falls. Similar curves to figures 7 and 8 can be generated for different 

chiller coolant in temperatures. 

6.4 Refrigerant Controls 

The same type and construction of expansion devices as the baseline, fixed orifice 

or a thermal expansion valve can be used in a secondary loop system. The device 

must be adjusted to reflect the R-152a refrigerant properties. In most cases, the 

refrigerant chiller out superheat setting for the secondary loop system is lower than 

the evaporator out superheat for the baseline system. As shown in figure 9, the low 

chiller superheat enables the chiller to operate at higher effectiveness. 

 Figure 8 Chiller Pressure Drop 

© 2007 Delphi Corporation, Inc. 
17

0.0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1.0

0 5 10 15 20 25 30 35 40

Chiller Out Super Heat (0 F)

Temperature Effectiveness 

εT=(Tcoolant in-Tcoolant out) / (Tcoolant in-Tsat at Pout)

2.8 5.6 8.3 11.1 13.9 16.7 19.4

(

0C)

6.5 Cooler 

The cooler is a secondary fluid coolant to air heat exchanger. Physically it replaces 

the traditional evaporator inside the HVAC module. 

. 

Figure 10 Typical Cooler 

Figure 9 Effect of Refrigerant Chiller Out Super 

 Heat on Chiller Effectiveness 

© 2007 Delphi Corporation, Inc. 

© 2007 Delphi Corporation, Inc. 
18

The secondary fluid (coolant) is chilled in the chiller and is pumped to the cooler 

where air is cooled to the desired temperature. Figure 11 illustrates the thermal 

performance of several coolers. Again, this heat exchanger has to be compact, 

lightweight and erosion / corrosion resistant. Such heat exchanger is shown in 

figure 9. It is aluminum brazed flat tube and center construction heat exchanger. 

350

370

390

410

430

450

470

490

510

530

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

Coolant Pressure Drop [psi]

Air Side Heat Transfer [BTU/min]

6.15

6.50

6.85

7.20

7.55

7.91

8.26

8.61

8.96

9.31

0 7 14 21 28 34 41 48 55 62 69 76 83 90 97

Coolant Pressure Drop [kPa]

Air Heat Transfer [kW]

58 mm (A) 58 mm (B) 73 mm 

47 mm (A) 47 mm (B) 55 mm (A) 

55 mm (B) 47 mm (C) 47 mm (D)

Target = > 414 BTU/min (7.27 kW), ∆P(Coolant) < 2.2 psig (15.2 kPa) 

Air In: 100°F x 40% RH (38°C ); Airflow Rate: 19.1 lbm/min (520 kg/hr) 

 Coolant In T: 30°F (-1.1°C); Coolant Flow Rate: 3, 4, 5, 6 gpm (11.4, 15.1, 18.9, 22.7 l/min) 

 Figure 11 Cooler Heat Transfer Performance

6.6 Pump 

The coolant pump is a centrifugal type with plastic housing and impeller. It is a 

variable speed and magnetically coupled to a brush-less motor. The pump picture 

and its performance are shown in figure 12. 

6.7 Secondary Fluid Selection (Coolant) 

Proper selection of the secondary fluid (coolant) is key to the optimum performance 

of a secondary loop system. It is necessary for the secondary fluid to be non-

© 2007 Delphi Corporation, Inc. 
19

flammable since it is circulated inside the vehicle passenger compartment. There 

are a variety of commercially developed, single phase heat transfer fluids which 

could be used as coolants. Glycols water mixtures, non-aqueous fluids such as 

synthetic organic fluids, Dowtherms, and Silicon oils are among the candidates. 

Because of their heat transfer degradation due their high viscosity at low 

temperatures, propylene glycol water solutions are mostly used in medium 

temperature applications above 0 0F (-18 0C). Ethylene glycol water solutions have 

better transport properties than their counterpart propylene solutions. Like any glycol 

solutions, increasing ethylene glycol concentration in water results in a performance 

penalty and an increase in pumping power. Therefore, the selection of the correct 

concentration for the secondary loop system applications is important. The lack of 

thermo-physical, safety, and automotive validation data coupled with the cost make 

the non-aqueous fluids not attractive at the moment. 

The range of operating temperatures of a cooler in an automotive secondary loop air 

conditioning system is between 32 0

F (0 0C) and 60 0F (15.5 0C), hence the 

Figure 12 Coolant Pump 
20

evaluation of secondary fluid thermo-physical properties have to be taken into 

consideration in this temperature range. Based on heat transfer properties, 

familiarity, wide automotive validation and applications, and the presence of 

inhibitors to prevent corrosion, engine coolant which is ethylene glycol based was 

selected for this project. Figure 13 show the burst and freeze lines of the various 

glycol water concentrations. 

 

 

A 32% ethylene glycol water solution exhibits adequate transport and thermophysical properties and provides sufficient heat transfer while protecting the coolant 

circuit components from freeze damages. The EG “burst” line depicted in the figure 

represents the “zero expansion” at low temperatures when the solution is not 

pumped and the EG line stands for the glycol percent concentration if the solution is 

pumped through the coolant circuit. Generally automotive air conditioning systems 

are controlled not to operate when the ambient temperature is below 40 to 45 0

 F 

(4 to 7 0C), therefore the 32% ethylene glycol concentration is ample enough to 

GLYCOLS

PG: Propylene 

EG: Ethylene 

 

TEG: Triethylene 

 

DPG: Dipropylene

DEG: Diethylene

Figure 13 Secondary Fluids (Coolants) 
21

protect the coolant components against very low ambient temperatures (~ -50 0F) 

when the air conditioning system is off. 

VII. PERFORMANCE COMPARISON 

The function of an A/C system is to quickly and efficiently maintain passenger’s 

thermal comfort irrespective of external ambient conditions. The most severe 

conditions a vehicle is subjected to are the soak and cool down and city traffic at 

high ambients. In this section, the performance of R-152a secondary loop system is 

compared to a conventional R-134a direct system using the same vehicle and tunnel 

under identical testing conditions. 

7.1 Test Description and Methodology 

The vehicle was first tested with its original equipment for baseline purposes. The 

same vehicle was outfitted with a secondary loop system with same or similar in size 

as the current R-134a components such as condenser. 

Table 1. Test Matrix

115 0F x 15% R.H. (46 0C)

* O.S.A. = Outside Air; Rec. Air = Re-circulated (Cabin Air); Hi Bl. = High Blower; F.C. = Full Cold

** R.L. = Road Load

(O.S.A., Hi Bl., F.C.)

Vent Mode 

(Rec. Air, Hi Bl., F.C.)

R.L. Stable Points: 30, 50, 70 (48, 80, & 113 kph)

 & Idle

R.L. Stable Points: 30, 50, 70 (48, 80, & 113 kph)

 & Idle

Ambient A/C setting Driving Conditions

104 0F x 40% R.H. (40 0C)

City Traffic Schedule

Soak & Cool Down; R.L.** Stable Points:

59 0F x 70% R.H. (15 0C)

95 0F x 40% R.H. (37 0C)

95 0F x 40% R.H. (37 0C)

Vent Mode 

Vent Mode 

(O.S.A / Rec. Air, Hi Bl., F.C.) *

Vent Mode 

Vent Mode 

(O.S.A., Hi Bl., F.C.)

Vent Mode 

(O.S.A., Hi Bl., F.C.)

(Rec. Air, Hi Bl., F.C.)

80 0F x 60% R.H. (27 0C)

30, 50, 70 (48, 80, & 113 kph) & Idle

R.L. Stable Points: 30, 50, 70 (48, 80, & 113 kph)

R.L. Stable Points: 30, 50, 70 (48, 80, & 113 kph)

 & Idle

 & Idle

Besides equipping the vehicle with standard instrumentation for climate control 

tunnel testing, the vehicle was also instrumented with several temperature sensors 

and pressure transducers to monitor all the relevant parameters and to extract 

maximum data information from both systems. 

The performance comparison of the two systems were made under identical test 

conditions. The tests were performed in an identical environment using the same 

vehicular thermal tunnel. Table 1 lists the test conditions under which both systems 

were evaluated. The test matrix was designed to provide enough performance data 

for comparison purposes. It is based on testing procedures used for validation and 

© 2007 Delphi Corporation, Inc. 
22

benchmarking of production and prototype air conditioning systems. The first set of 

tests shown on the table were dedicated to high loads (high ambient) while the rest 

of the test matrix was devoted to capacity control for the purpose of reducing the 

energy usage by the secondary loop system. 

7.2 Performance under High Ambient Conditions 

In general, passenger comfort is related to the air conditioning discharge and cabin 

temperatures coupled with airflow and its circulation inside the passenger 

compartment. Figure 14 shows the dynamic changes of the discharge and cabin 

temperatures under ambient conditions of 115 0

F (46 0C) x 15% relative humidity.

These conditions are representative of the Phoenix area summer climate. The 

driving cycle consisted of a soak, followed by a city traffic schedule in re-circulated 

mode and finally an idle in outside air mode. 

As seen in the graph, at various driving conditions the behavior of the vent discharge 

and cabin temperature variations between the secondary loop system and the 

Performance Comparison @ 46 0C x 15% R.H. Ambient 

0

10

20

30

40

50

60

70

0 10 20 30 40 50 60 70 80

Time (minutes)

Temperature ( C)

40 kph 

Recirc.

Idle 

OSA

100 kph 

OSA

Idle

Recirc.

100 kph 

Recirc.

50 kph 

Recirc.

City Traffic 

Recirc.

Blue: Baseline (R-134a) Red: R-152a S.L.

 Average Cabin

 Average Vent

Figure 14 Performance Comparison at High Loads

© 2007 Delphi Corporation, Inc. 
23

baseline production system are almost identical through the driving cycle regardless 

of supplied air, being in outside air (fresh air) or re-circulated air (cabin air). Figure 

15 presents the cooling performance comparison of the two systems at 104 0

F (40 

0C) x 40% relative humidity under soak and cool conditions, and stable driving 

conditions. Again, the discharge and cabin temperatures are almost identical 

through the driving cycle. 

Performance Comparison @ 40 0C x 40% R.H. Ambient 

0

10

20

30

40

50

60

70

0 20 40 60 80 100 120 140 160

Time (minutes)

Temperature (

0C)

50 kph 

Recirc

Idle 

OSA

110 kph 

OSA

Idle

Recirc

80 kph 

OSA

50 kph 

OSA

80 kph 

Recirc

Blue: Baseline (R-134a) Red: R-152a S.L.

 Average Cabin

 Average Vent

The secondary loop system response time during soak and cool conditions 

sometimes called “thermal lag” was originally a concern because of the thermal 

capacitance associated with the coolant. However, as shown in figures 14 and 15, 

both systems were soaked to the same temperature, yet the response time delay 

(thermal lag) of the secondary loop system is negligible. The thermal lag concerns 

are emanated from thoughts that the coolant temperature would match the soak 

temperature (cabin temperature) when the vehicle is soaking in the sun. However, 

data has shown that the coolant temperature stays at ambient temperature since the 

whole system including the coolant reservoir is under-hood and not subjected to the 

effect of the sun load. Over the years, the results of numerous subjective evaluations 

of the secondary loop carried out under high ambient conditions concluded that the 

thermal lag is not an issue. The thermal lag is certainly measurable however not 

Figure 15 Performance Comparison @ High Loads 

© 2007 Delphi Corporation, Inc. 
24

perceptible. Some of the recent subjective evaluation is described in the following 

section. 

7.3 Road Performance under High Loads

One of the four demonstration vehicles shown at the 2007 Alternate Refrigerant 

Symposium was equipped with an R-152a secondary loop system. 

2007 SAE Phoenix Alternate Refrigerant System Symposium

Vent Temperature at Average Ambient: 107 0F (41.7 0C) x 24% R.H.

40

45

50

55

60

65

70

1

Driving Schedule

(

0

 F)

R-134a Baseline

DP-1

AC-1

R-152a Secondary Loop

C.T.* (1) C.T. (2) C.T. (3) C.T. (4) 30 min. 35 min. 40 min. 45 min.

 4.4

 10

12.8

21.1

(

0C)

18.3

15.5

7.2

* C.T.= City Traffic

The forum [6] was held in Scottsdale, Arizona and was represented by almost 270 

attendees. The participants came mainly from the mobile air conditioning industry. 

Besides discussing the progress of alternate refrigerant technologies, the forum 

covered the comfort evaluation of the demonstration vehicles. The evaluation 

Figure 16 Road Performance at High Loads 
25

consisted of a subjective comfort rating coupled with air vent and cabin temperature 

readings at each phase of the driving cycle. 

2007 SAE Phoenix Alternate Refrigerant System Symposium

Cabin Temperature at Average Ambient: 107 0F (41.7 0C) x 24% R.H.

50

60

70

80

90

100

1

Driving Schedule

(

0

 F)

 10

 21.1

37.8

(

0C)

32.2

26.7

15.5

C.T.* (1) C.T. (2) C.T. (3) C.T. (4) 30 min. 35 min. 40 min. 45 min.

R-134a Baseline, Soak = 145.3 0F (62.9 0C) 

DP-1, Soak = 142.7 0F (61.5 0C) 

AC-1, Soak = 143.8 0F (62.1 0C) 

R-152a Secondary Loop, Soak = 146.8 0F (63.8 0C) 

* C.T.= City Traffic

The demonstration vehicles were identical (2007 Opel Astra) including exterior and 

interior colors. The assessors were not informed on what type of refrigerant or 

system is in each vehicle (blind evaluation). The rides were random and blind for the 

purpose of eliminating any bias for or against a typical refrigerant or system. Figures 

16 and 17 display the objective evaluation, which is represented by the 

measurements of the vent and cabin temperatures of the R-134a baseline, DP-1 

refrigerant system, AC-1 refrigerant system, and R-152a secondary loop system. 

This objective ratings show that the baseline and the R-152a secondary loop system 

are similar in performance and are better performers than DP-1 and AC-1 systems. 

Figure 17 Road Performance at High Loads 
26

2007 SAE Phoenix Alternate Refrigerant System Symposium

Comfort Rating at Average Ambient: 107 0F (41.7 0C) x 24% R.H.

0

1

2

3

4

5

6

1

Driving Schedule

Rating

C.T.* (1) C.T. (2) C.T. (3) C.T. (4) 30 min. 35 min. 40 min. 45 min.

R-134a Baseline

DP-1

AC-1

R-152a Secondary Loop

5 --- Comfort

4 --- Slightly Warm

3 --- Warm

2 --- Hot

* C.T.= City Traffic

Comfort

The results of the subjective comfort rating are shown in figure 18. They follow 

similar patterns as the objective evaluation. Over all the performance of the R-152a 

secondary loop system matches the baseline system. 

7.4 Performance under Medium to Mild Ambient Conditions 

Mobile air conditioning systems are generally designed and sized for maximum 

loads which in most cases are soak and cool downs under extreme environmental 

conditions. Under medium ambient conditions such as 95 0

F (35 0C) x 40% relative 

humidity and mild ambient conditions such as 80 0F (27 0C) x 60% relative humidity, 

overcooling occurs in nearly all vehicles resulting in a reheating process by 

tempering with the temperature door, sometimes called blend door to maintain 

comfort. This occurrence is sometimes called “series reheat” in the industry. In order 

to reduce the series reheat, today’s systems require sophisticated and expensive 

controls to operate 

Figure 18 Subjective Comfort Evaluation at High Loads 
27

 Discharge Temperature Comparison (10 min. Point) 

@ 95 0

F x 40% R.H. (35 0C) Ambient, High Blower, OSA 

0

10

20

30

40

50

60

70

1

(

0

F)

Baseline S.L. with 4 gpm (15.14 

lpm)

S.L. with 3.5 gpm (13.25 lpm)

 - 6.6

 4.4

15.5

21.1

(

0C)

30 mph (48 kph) 50 mph (80 kph) 70 mph (113 kph) Idle

Cabin Temperature Comparison (10 min. Point) 

@ 95 0F x 40% R.H. (35 0C) High Blower, OSA 

60

62

64

66

68

70

72

74

76

78

80

1

(

0

F)

30 mph (48 kph) 50 mph (80 kph) 70 mph (113 kph) Idle

 15.5

17.8

20.0

22.2

24.4

26.7

(

0C)

Baseline S.L. with 4 gpm (15.14 lpm)

S.L. with 3.5 gpm (13.25 lpm)

 Figure 19a Overcooling Reduction 

Figure 19b Overcooling Reduction 

© 2007 Delphi Corporation, Inc. 

© 2007 Delphi Corporation, Inc. 
28

the evaporator at higher pressures. In a conventional system, operating the 

evaporator at higher pressures elevates the evaporator air outlet temperature thus 

minimizing the reheating process and in turn reducing the compressor shaft power. 

The secondary loop system offers this benefit of overcooling prevention by simply 

varying the coolant flow using the speed of the pump, which in turn, reduces the 

compressor shaft power. This technique of “series reheat” reduction is not 

complicated and is less costly. 

 Discharge Temperature Comparison (10 min. Point) 

@ 80 0

F x 60% R.H. (27 0C) Ambient & Blower Setting 3

0.0

10.0

20.0

30.0

40.0

50.0

60.0

70.0

1

(

0

F)

Baseline S.L. with 3 gpm (11.35 lpm) S.L. with 2.5 gpm (9.46 lpm)

-12.2

- 6.6

- 1.1

4.4

10.0

15.5

21.1

(

0C)

30 mph (48 kph) 50 mph (80 kph) 70 mph (113 kph) Idle

Figures 19a and 19b depict a typical wind tunnel test data that represents the 

cooling performance comparison of the baseline and the secondary loop system at 

95 0

F (35 0C) x 40% relative humidity. The tests were conducted at vehicle speeds of 

30, 50, 70 mph (48, 80, 113 kph) and idle. In the data shown in figures 19a and 19b, 

the blower was set to high position while the air supply was dialed in outside air 

mode (fresh air). The airflow rate in outside air mode at high blower position is 

around 260 ft3/min. (7.36 m3/min.) for both systems. In re-circulated air mode at 

blower setting 3 which represents high medium blower, on the average the airflow 

rate for the secondary loop system is higher than the baseline. The average amount 

of air flow rate in re-circulating mode for the secondary loop system at blower setting 

3 is 216 ft3

/min. (6.12 m3/min.) while the baseline air flow rate is 205 ft3/min. (5.80 

m

3

/min.). Table 2 shows the airflow rates at various HVAC blower settings. The 

Figure 20a Overcooling Reduction 

© 2007 Delphi Corporation, Inc. 
29

variation among the airflow rates between the baseline and the secondary is mainly 

due to the depth difference between the baseline evaporator and the secondary loop 

cooler. 

 Cabin Temperature Comparison (10 min. Point) 

@ 80 0

F x 60% R.H. (27 0C) Ambient & Blower Setting 3

50.0

55.0

60.0

65.0

70.0

75.0

1

(

0

F)

 10.0

12.8

15.5

18.3

21.1

23.9

(

0C)

Baseline S.L. with 3 gpm (11.35 lpm)

S.L. with 2.5 gpm (9.46 lpm)

30 mph (48 kph) 70 mph (113 kph) Idle 50 mph (80 kph)

As shown in graph 19a, when the coolant pump is run at full capacity (4 gpm / 15.14 

lpm), on the average the secondary loop vent discharge temperatures are lower than 

the baseline temperatures. Dialing down the coolant flow rate to 3.5 gpm, (13.25 

lpm) raised the vent temperatures up to par with baseline or higher, thus avoiding 

overcooling. After long driving periods, in current mobile air conditioning systems 

without series reheat reduction option, overcooling may take place and most vehicle 

operators would dial up the temperature door (blend door) for reheating purpose. 

However, this maneuver consists mainly of redirecting a certain percentage of the 

Figure 20b Overcooling Reduction 

© 2007 Delphi Corporation, Inc. 

© 2007 Delphi Corporation, Inc. 
30

cold air toward the heater core for warmth and later on blends it with the outgoing 

vent air. This process does not reduce or influence the load on the system hence it 

does not diminish the compressor work. 

Instead of relying on temperature door (blend door) movements for reheat, the 

secondary loop system uses the coolant flow variation as the controlling parameter. 

This reheating process is accomplished by adjusting the coolant pump speed 

(decreasing coolant flow) which in turn lowers the load on the system and hence 

reducing the compressor shaft work. The compressor work reduction is achieved by 

cycling off a fixed displacement compressor or de-stroking a variable displacement 

compressor. Figure 19a illustrates the reheating process reduction in outside air 

mode by turning down the coolant from 4 gpm (15.14 lpm) to 3.5 gpm (13.25 lpm). 

Conventionally in a baseline, system with the reheating option or in a secondary loop 

system with modulating coolant flow for reheating purposes the passenger comfort is 

always respected and not compromised. 

 Discharge Temperature Comparison (10 min. Point) 

@ 80 0

F x 60% R.H. (27 0C) Ambient & Blower Setting 2

0.0

10.0

20.0

30.0

40.0

50.0

60.0

1

(

0

F)

Baseline S.L. with 3 gpm (11.35 lpm)

S.L. with 2.5 gpm (9.46 S.L. with 2.0 gpm (7.57 

-12.2

- 6.6

- 1.1

4.4

10.0

15.5

(

0C)

30 mph (48 kph) 50 mph (80 kph) 70 mph (113 kph) Idle

In today’s air conditioning systems, under mild ambient conditions such as 80 0F (27 

0C) x 60% relative humidity and low ambient conditions, overcooling very frequently 

occurs and in most cases to maintain comfort, reheating using temperature door 

(blend door) levels and/or airflow reduction is needed. 

Figure 21a Overcooling Reduction 

© 2007 Delphi Corporation, Inc. 
31

 Cabin Temperature Comparison (10 min. Point) 

@ 80 0

F x 60% R.H. (27 0C) Ambient & Blower Setting 2

50.0

55.0

60.0

65.0

70.0

75.0

1

(

0

F)

Baseline S.L. with 3 gpm (11.35 lpm)

S.L. with 2.5 gpm (9.46 

lpm)

S.L. with 2.0 gpm (7.57 lpm)

 10.0

12.8

15.5

18.3

21.1

23.9

(

0C)

30 mph (48 kph) 70 mph (113 kph) 50 mph (80 kph) Idle

Figures 20a through 21b show the cooling performance comparison of the baseline 

and the secondary loop at 80 0F (27 0C) x 60% relative humidity with air supply set in 

outside air mode (fresh air). At this mild ambient and below, most vehicle operators 

prefer fresh air versus cabin air unless somehow the outside fresh air is 

contaminated. As previously stated, the tests were conducted at vehicle speeds of 

30, 50, 70 mph (48, 80, 113 kph) and idle. 

Figures 20a and 20b describe the discharge and cabin temperatures under 

moderate loads portrayed by high medium blower (setting 3) and outside air 

conditions. Instead of operating the pump at maximum coolant flow, which is in this 

case over 4 gpm (15.14 lpm), the pump is operated at 3 or 2.5 gpm (11.35 or 9.46 

lpm). The graphs show that overcooling reduction can be achieved by operating the 

pump at 2.5 gpm (9.46 lpm) or below without compromising the comfort. Under 

lower loads where the blower is set at medium (setting 2) or lower, the coolant is cut 

down further to 2 gpm (7.57 lpm) or lower, therefore reducing the load on the 

primary system and consequently the compressor shaft work is reduced. The load 

on the primary system, which is mainly the compressor power, is brought down by 

reducing the overcooling depending on the desired comfort. Figures 21a and 21b 

present an array of load reduction on the system by means of low coolant flow rates 

when the A/C demand is not high (low load). 

Figure 21b Overcooling Reduction 

© 2007 Delphi Corporation, Inc. 
32

VIII. ENERGY COMPARISON 

The coefficient of performance (COP) comparison was carried under 104 0F (40 0C) 

x 40%, 95 0F (37 0C) x 40%, 80 0F (27 0C) x 60% relative humidity. The COP 

discussed in this section is the thermal COP, which is defined as the ratio of the 

refrigerant cooling over the compressor power. 

 

Where: 

Heat Exchanger represents the evaporator and chiller for baseline and secondary 

loop respectively, and h is the refrigerant enthalpy Btu / lbm (kJ/kg) calculated at 

measured inlet and outlet pressure & temperature of each component. 

8.1 Energy Comparison at Full Capacity 

As depicted in figure 22, at the same capacity when compared to the baseline, the 

secondary loop system has on the average a lower COP. The outcome was 

COP Comparison @ 40 0C x 40% R.H. Ambient 

0

0.5

1

1.5

2

2.5

1 2 3 4 5 6 7

Driving Conditions

COP

Blue: Baseline (R-134a) Red: R-152a S.L.

50 kph

(OSA)

80 kph

(OSA)

110 kph

(OSA)

Idle (10) 

(OSA)

Idle (21) 

(OSA)

Idle (10) 

(Recir.)

Idle (21) 

(Recir.)

- 5 % + 0.3 % - 4 % - 1.5 % - 2 % - 1 % - 4 %

h out (Heat Exchanger) - h in (Heat Exchanger) 

 h out (Compressor) - h in (Compressor) 

COP =

Figure 22 Coefficient of Performance 

© 2007 Delphi Corporation, Inc. 
33

 Coefficient of Performance (COP) Comparison (10 min. Point) 

@ 95 0F x 40% R.H. (35 0C) Ambient, High Blower & OSA 

1

1.2

1.4

1.6

1.8

2

2.2

2.4

1

COP

30 mph (48 kph) 50 mph (80 kph) 70 mph (113 kph) Idle

Baseline S.L. with 4 gpm (15.14 lpm)

S.L. with 3.5 gpm (13.25 lpm)

expected and was proven before by the author [4] in his study of secondary loop. 

The COP percentage difference between the baseline and the secondary loop is 

lower in this investigation than the previous study. This is attributed to R-152a being 

a better refrigerant than R-134a. As we all know the coefficient of performance 

(COP) is the ratio of cooling capacity versus the energy usage of a system. In more 

simple terms, it is what you get in cooling benefit or comfort versus (useful effect) 

over what you pay for (energy that must be purchased) to get the benefit. 

Technically the COP cannot be labeled as efficiency since it is higher than one, 

however it is a measure of how environmentally or economically efficient the cooling 

(comfort) is attained. The COP is not a measure of energy usage, it is more like how 

cost-effective or otherwise the energy is paid for through fuel economy or emission. 

8.2 Energy Comparison at Controlled Capacity 

Figures 23 through 24b compare the coefficient of performances (COPs) of the 

baseline system with several control capacity options of the secondary loop system 

at medium and low loads. The graphs clearly show that the COP gap between the 

baseline and secondary loop is widening as the load is decreased. This behavior 

was also observed by the SAE report on improved efficiency of mobile air 

conditioning (IMAC) [8]. The lower COPs for the secondary loop system are 

Figure 23 Coefficient of Performance (COP) Comparison 

© 2007 Delphi Corporation, Inc. 
34

attributed to two factors, the lower compressor efficiency at reduced compressor 

strokes and compressor operation with R-152a instead of R-134a. 

 Coefficient of Performance (COP) Comparison (10 min. Point) 

@ 80 0F x 60% R.H. (27 0C) Ambient & Blower Setting 3

0

0.5

1

1.5

2

2.5

3

3.5

4

1

COP

30 mph (48 kph) 50 mph (80 kph) 70 mph (113 kph) Idle

Baseline S.L. with 3 gpm (11.35 lpm) S.L. with 2.5 gpm (9.46 lpm)

 Coefficient of Performance (COP) Comparison (10 min. Point) 

@ 80 0F x 60% R.H. (27 0C) Ambient & Blower Setting 2

0

0.5

1

1.5

2

2.5

3

3.5

4

1

COP

30 mph (48 kph) 50 mph (80 kph) 70 mph (113 kph) Idle

Baseline S.L. with 3 gpm (11.35 lpm)

S.L. with 2.5 gpm (9.46 lpm)

S.L. with 2.0 gpm (7.57 lpm)

Figure 24a Coefficient of Performance (COP) Comparison 

Figure 24b Coefficient of Performance (COP) Comparison 

© 2007 Delphi Corporation, Inc. 

© 2007 Delphi Corporation, Inc. 
35

As in IMAC study, the energy usage or compressor power is a better metric than 

COP to evaluate fuel usage and emissions. The cooling capacity is directly 

proportional to the cooler (evaporator) air out temperature. The capacity reduction is 

accomplished by increasing the cooler (evaporator) air out temperature while 

keeping the comfort at the desired level. For the secondary loop system, the 

increase of cooler air out temperature is achieved by means of reducing the coolant 

flow rate. 

Secondary Loop Capacity Control Analysis

 Blower (Blwr) Settings: 3 and 2, Outside Air Mode 

Ambient: 80 0C (27 0C) x 60% R.H.,

0.0

0.5

1.0

1.5

2.0

2.5

4 5 6 7 8 9 10 11 12

Cooler (Evaporator) Air Out 

Compressor Power (kW)

R134a at

Blwr3

SL R152a at

Blwr3

R134a at

Blwr2

SL R152a at

Blwr2

Minimum Air Temp. (Maximun Cooling)

No Capacity Control (No Series Reheat)

Road Load: Compressor in De-Stroked Position

(

0C)

41 42.8 44.6 46.4 48.2 50 51.8

(

0

F)

Maximum Air Temp.(Minmum Cooling)

Full Capacity Control (Full Series Reheat)

Capacity Control Slope 

Path

Figures 25a and 25b describe the compressor power variation with the cooler 

(evaporator) air out temperature. The data points on the left and the right of the 

graphs illustrate the maximum cooling capacity (lowest cooler / evaporator air out) 

and the minimum cooling capacity (highest cooler / evaporator air out) respectively. 

It is evident from the graphs, at identical cooling capacities; the secondary loop 

system compressor uses more power than the baseline compressor. Most of the 

time, at medium to low loads, operating at maximum cooling capacity is not 

necessary. It initiates overcooling and as mentioned previously a reheating is 

needed to maintain comfort. Operating the secondary loop system below the blue 

Figure 25a Compressor Power versus Capacity at 

 Rod Load Conditions 

© 2007 Delphi Corporation, Inc. 
36

dashed lines shown on the graphs will reduce overcooling. The capacity is controlled 

to match the desired comfort without relying on reheating. As shown in figure 25a at 

road load, if the cooler air out temperature is controlled to 46 0

F (8 0C) or above the 

secondary loop system with controlled capacity will use less compressor shaft power 

than the baseline. At idle, the control temperature is 49 0F (9.5 0C). As seen on the 

plots, the secondary loop system compressor power can be maintained at equal or 

lower values of the baseline compressor power by means of controlling the cooling 

capacity. 

1.0

1.2

1.4

1.6

1.8

2.0

2.2

2.4

6 7 8 9 10 11 12

Compressor Power (kW)

R134a at Blwr3

SL R152a at

Blwr3

R134a at Blwr2

SL R152a at

Blwr2

IDLE: Compressot at Full Stroke Position

Cooler (Evaporator) Air Out 

(

0C)

44.6 46.4 48.2 50 51.8 (

0

F)

Secondary Loop Capacity Control Analysis

Blower (Blwr) Settings: 3 and 2, Outside Air Mode

Ambient: 80 0C (27 0C) x 60% R.H.

Capacity Control Slope 

Path

Minimum Air Temp. (Maximun Cooling)

No Capacity Control (No Series Reheat)

Maximum Air Temp.(Minmum Cooling)

Full Capacity Control (Full Series Reheat)

As mentioned before, the capacity control is achieved through modulating the 

coolant flow rate. At medium to low loads, the secondary loop system capacity 

control is an effective way of reducing the compressor shaft power. The next 

section addresses the benefit in fuel consumption generated through the reduction in 

compressor shaft power. 

Figure 25b Compressor Power versus Capacity at 

Idle Conditions

© 2007 Delphi Corporation, Inc. 
37

8.3 Fuel Consumption Analysis 

The objective of this section is to estimate the air conditioning fuel use of a baseline 

R-134a and a secondary loop R-152a for various driving schedules. National 

Renewal Energy Laboratory (NREL), because of their experience and expertise in 

this field has cooperated with Delphi to carry out the fuel usage estimation. John 

Rugh, a senior engineer at NREL, has championed the analysis. 

0 2 0 0 4 0 0 6 0 0 8 0 0 1 0 0 0 1 2 0 0

0

2 0

4 0

6 0

8 0

C Y C _ N E D C

t im e (s e c )

speed (mph)

k e y o n

s p e e d

e le va t io n

Figure 26a Speed Correction Schedule (SC03) 

Figure 26b New European Driving Cycle (NEDC) 
38

Figure 26c Highway Fuel Economy Test 

Figure 27 Sample of Inputs to the “Advisor” Software 
39

The inputs to the model consisted of the compressor power versus speed generated 

from vehicle tunnel tests at ambient of 80 0

F (27 0C) x 60% relative humidity, the 

vehicle characteristics such as mass and engine type, and the drive cycle data. The 

driving schedules simulated in this analysis are the New European Driving Cycle 

(NEDC), the Speed Correction Cycle (SC03), and Highway Fuel Economy Test 

(HWFET). These driving schedules depicted in figure 26a through 26c are run with 

the A/C off and A/C on. The vehicle model is built using “Advisor Software” with 

inputs such as coefficient of drag, frontal area, mechanical accessory load, and 

rolling resistance coefficient and engine characteristics. A sample of these inputs to 

the software is shown in figure 27. The vehicle and A/C data are shown in figure 28. 

Coefficient of Drag: 0.32; Frontal Area: 2.11 m2 (22.7 ft2) Mass: 1263 kg 

(2784 lbm), R134a 1270 kg (2799 lbm), SL152a A/C off Mechanical 

Accessory Load = 500 W 

Rolling Resistance Coefficient = 0.009 

Engine – 63 kW 1.9 L Saturn Engine Scaled to 75 kW 

The vehicle A/C data was taken under the ambient of 80 0F (27 0C) x 60% relative 

humidity, vent mode, full cold and the blower setting was at position 2 which 

Figure 28 Vehicle and A/C Data 

© 2007 Delphi Corporation, Inc. 
40

corresponds to low medium blower (~ 6 volts). The coolant pump power of 15 Watts 

was taken into account in the analysis. 

Figure 29 shows the total fuel consumption per 100 km of the vehicle with the air 

conditioning off, with R-134a system on, and with the R-152a secondary loop system 

with control capacity on. As expected the total fuel consumption with A/C off at each 

driving cycle is the lowest, however with A/C on, the consumption is less with the 

control capacity secondary loop system than the baseline R-134a system. 

 

Figure 30 shows the air conditioning fuel usage per 100 km comparison between the 

baseline R-134a system and the R-152a secondary loop system with controlled 

capacity. Depending on the driving cycle, the advantage of the secondary loop with 

controlled capacity ranges from 13 to 16%. This fuel saving can be achieved by 

reducing overcooling and eliminating the unnecessary reheating process to reach 

comfort. As stated before, averting the overcooling would ultimately translate into 

saving energy, hence less fuel use and less emission. 

8.0

8.3

6.3

8.9

9.3

6.7

8.8

9.2

6.6

0

1

2

3

4

5

6

7

8

9

10

SCO3 NEDC HEFET

Fuel Consumption (L/100 km)

A/C off

R134a

SL 152a w/ CC

Figure 29 Overall Vehicle Fuel Consumption with A/C Off and 

A/C on 
41

IX. SECONDARY LOOP SYSTEM DESIGN GUIDELINES 

For comparison purpose, it is recommended to baseline the vehicle before equipping 

it with Secondary Loop system. 

9.1 Overall System 

• To meet maximum performance requirement, it is recommended to 

use higher displacement (10 – 15%)compressor 

 Ex: Assuming same speed & isentropic efficiency for both 

 compressors and using saturation properties of the 

 refrigerants, @ 5, 10, 15 & 27 0C compressor in 

 temperatures, the increase in displacement is around 11.5% 

• Verify the desiccant type (XH-9 is required) 

• Re-set engine cooling fan control to match R-152a properties 

0.0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1.0

1.1

SCO3 NEDC HEFET

A/C Fuel Use (L/100 km)

R134a

SL R152a w/ CC

-13%

-13%

-16%

0.0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1.0

1.1

SCO3 NEDC HEFET

A/C Fuel Use (L/100 km)

R134a

SL R152a w/ CC

-13%

-13%

-16%

Figure 30 A/C Fuel Consumption 
42

 (recalibrated high side pressure transducer for R-152a. It is usually 

 located around condenser) 

• Check front end air flow (minimize air re-circulation) 

• The TXV needs to be reset to match R-152a properties and the 

secondary loop applications 

• Cooler coolant in temperature (instead of cooler air off) is 

recommended for freeze control 

• Capacity controller 

• Insulate refrigerant in line to chiller 

• Insulate chiller 

• Depending on the length, coolant lines could be insulated 

• Coolant circuit: 

 Pump located at lowest point of circuit 

 Reservoir fill cap located at highest point (open to atmosphere for 

coolant circuit de-gassing) 

 Chiller on suction side of pump (pull coolant flow through 

 chiller for better coolant distribution in the chiller) 

 Minimum reservoir size for faster cool down 

 Dual reservoir is recommended for idle - stop 

 Coolant mix: 32% Glycol / 70% water 

More details on secondary loop system design guidelines are described in the 

design improvements section. 

9.2 Potential Design Improvements 

As with all design efforts, optimization is a series of trade-offs between various 

features or metrics. In the design of a Secondary Loop System, some key items to 

consider for optimization of both performance and minimum energy use of the 

design are: 

• System Mass

• Heat Exchanger Effectiveness

• Component Package Size

• Compressor Efficiency

• Control Strategies for minimum Fuel Use

• “Cold” Storage for Energy Savings
43

Each of these features will be summarized in order to give options for performance 

and/or a reduction in the energy needed for operation. 

9.2.1 System Mass 

For the Astra demonstration vehicle, the additional mass added to support a 

Secondary loop system is summarized in table 3. 

Component Added mass Comment 

Chiller 1.8 kg (including fittings) Based demonstration vehicle content 

Cooler Similar to evaporator Based demonstration vehicle content 

Pump 0.43 kg 

Coolant 3.2 kg Function of project requirements 

Total Mass 5.45 kg Function of project requirements 

As is shown, the chiller and coolant make up the majority of the added mass of the 

system. As an improvement, a reduction in mass can be achieved by optimizing the 

designs. Mass of the chiller depends on several factors: its required size to meet the 

heat transfer requirements in the vehicle, the necessary material gage to meet field 

reliability targets, its mounting features in the vehicle, its plumbing to interface with 

primary refrigerant and the coolant circuits. The amount of insulation added to the 

chiller assembly to minimize the environmental heat gain could add to the mass of 

the chiller assembly. Guidelines to select a chiller design based on these 

characteristics are: 

1. Select the chiller size to meet the cooling requirements, as well as, 

balance the heat transfer and the chiller effectiveness versus other 

components in the system, particularly the pressure drop (coolant and 

refrigerant). An optimum chiller design will depend on the required heat 

exchanger surface area, the overall heat transfer coefficient, and the 

refrigerant / coolant flow rates. 

2. The gage of the material selected must be consistent with the 

requirement of the SAE J639 specification, the material gage must 

meet the manufacturing, and durability needs as specified by each 

OEM. 

3. Chiller mounting features such as brackets and fittings have an impact 

on the mass of the chiller. Directionally, it is preferred to encase the 

metallic heat exchanger in a plastic enclosure and use plastic features 

to mount the assembly to the vehicle. 

4. The required plumbing to interface with the primary and secondary 

fluids will depend on the type of refrigerant control and the position of 

Table 3 System Mass 

© 2007 Delphi Corporation, Inc. 
44

the inlet and outlet fittings. The pass arrangement and the type of 

fittings used will influence the plumbing layout. 

5. The addition of insulation to a chiller assembly is suggested to 

minimize heat loss; however, the environment where the chiller is 

located in the vehicle will determine the quality and the quantity of 

insulation. 

The major source of additional mass is the amount of coolant used in the secondary 

fluid (glycol/water) loop. 0.75 gallon (3.2 liters) of a 30% glycol / 70% water mix was 

used in the Astra demonstration vehicle. 

In any given application, there is a minimum amount of coolant required to fill the 

system volume while maintaining system operation. Additional coolant can be added 

to provide extra “cold” storage. For the Astra vehicle, the minimum amount was 

approximately 0.77 gallon (2.9 liters) was required to fill the system. As a reserve for 

minimum cold storage, an additional 0.08 gallon (0.3 liters) was added via an in-line 

reservoir. With this amount of coolant in the system, the Astra system was able to 

maintain vehicle comfort up to three times longer then the baseline R-134a 

production system. This level of performance was measured in the hot ambient 

conditions of Phoenix, AZ and it is shown in figure 31. The longer the desired A/C off 

time, the greater the coolant storage volume. 

 

 Figure 31 High Ambient System Warm-up with 

 Cold Storage 

R-152a Secondary Loop Re-Heat Evaluation

0

5

10

15

20

25

30

35

0 20 40 60 80 100 120 140

Time [sec]

AVG Vent Temperature [°C]

2nd Loop - Pumps Off

2nd Loop - Pumps @ 1.5 gpm

2nd Loop - Pumps @ Full

R-134a Baseline #1

R-134a Baseline #2

R-134a Baseline #3

TEST CONDITIONS

Ambient Temperature: 40~43°C

Vehicle Speed: 80 kph to Idle

HVAC Mode: Vent, Full Cold, Recirc

Coolant Pumps @ 80 kph: Full Power

Coolant Pumps @ Idle: 4, 1.5, 0 gpm

Comfort time (with compressor off) can be 

more than 3x longer with SLS 

Initial Temperature Offset Correction

R-152a Secondary Loop Re-Heat Evaluation

0

5

10

15

20

25

30

35

0 20 40 60 80 100 120 140

Time [sec]

AVG Vent Temperature [°C]

2nd Loop - Pumps Off

2nd Loop - Pumps @ 1.5 gpm

2nd Loop - Pumps @ Full

R-134a Baseline #1

R-134a Baseline #2

R-134a Baseline #3

TEST CONDITIONS

Ambient Temperature: 40~43°C

Vehicle Speed: 80 kph to Idle

HVAC Mode: Vent, Full Cold, Recirc

Coolant Pumps @ 80 kph: Full Power

Coolant Pumps @ Idle: 4, 1.5, 0 gpm

Comfort time (with compressor off) can be 

more than 3x longer with SLS 

Initial Temperature Offset Correction

© 2007 Delphi Corporation, Inc. 
45

9.2.2 Heat Exchanger Effectiveness 

Although the secondary loop system has many benefits as we have discussed, the 

coolant circuit adds a “heat transfer resistance” that must be overcome to provide 

the necessary temperature difference. One common measure of how well heat is 

transferred in a given heat exchanger is its effectiveness. A heat exchanger 

effectiveness is defined as the ratio of the actual heat transfer over the theoretical 

maximum heat transfer in the heat exchanger. Since the coolant is a common fluid in 

the chiller and cooler, the two heat exchangers are considered coupled. London and 

Kays [9] have established an overall effectiveness of such coupled heat exchangers. 

Figure 32 shows the relationship between the power required by a secondary loop 

system as compared to an R-134a direct expansion system and the effectiveness of 

the heat exchangers at a fixed system load and range of evaporating and 

condensing temperatures. It is based on the Astra vehicle data with R-152a 

secondary loop system data. 

It is important to note that by increasing the combined effectiveness of the 

cooler/chiller as compared to the direct expansion system evaporator effectiveness 

reduces the power required. The ratio for equivalent power is 0.9 or less. The fact 

that this ratio is less than 1 is a result of using the R-152a refrigerant. Therefore, as 

Compressor Power Ratio for SL R152a versus Evaporator Effectiveness Ratio

Assumes: Same Compressor Speed, Isentropic Efficiency, and Condensing Temperature

0.80

0.85

0.90

0.95

1.00

1.05

1.10

1.15

1.20

1.25

1.30

1.35

1.40

0.80 0.85 0.90 0.95 1.00

EFF Cooler+Chiller/EFF Evap

Comp Power SL/Comp Power Base

Road Load Condensing T Idle Condensing T

Evap T3

(WR152a-SL/WR134a)=(אhevap-R134a*אhisen-R152a*אisen-R134a)/(אhevap-R152a*אhisen-R134a*אisen-R152a)

Evap T1

Evap T2

Opel Astra 

EFF Ratio 

R152a Direct Expansion 

System vs. R134a

Lower Energy Use

Compressor Power Ratio for SL R152a versus Evaporator Effectiveness Ratio

Assumes: Same Compressor Speed, Isentropic Efficiency, and Condensing Temperature

0.80

0.85

0.90

0.95

1.00

1.05

1.10

1.15

1.20

1.25

1.30

1.35

1.40

0.80 0.85 0.90 0.95 1.00

EFF Cooler+Chiller/EFF Evap

Comp Power SL/Comp Power Base

Road Load Condensing T Idle Condensing T

Evap T3

(WR152a-SL/WR134a)=(אhevap-R134a*אhisen-R152a*אisen-R134a)/(אhevap-R152a*אhisen-R134a*אisen-R152a)

Evap T1

Evap T2

Opel Astra 

EFF Ratio 

R152a Direct Expansion 

System vs. R134a

Lower Energy Use

Figure 32 Compressor Power versus Effectiveness 

© 2007 Delphi Corporation, Inc. 
46

a guideline, designs that maximize the combined heat exchanger effectiveness will 

result in lower energy requirements for the secondary loop system. 

9.2.3 Component Package Size 

Packaging the components that make up a secondary loop system is one of the 

challenges in applying this technology to production vehicle applications. An 

advantage of the secondary loop system components is that they do not require a 

specific in-vehicle location to perform the required function. For example, power train 

cooling heat exchangers are located forward in the vehicle and require front-end 

airflow to operate effectively. 

Guidelines to efficient packaging of the components: 

• Locate the chiller away from significant heat sources such as turbocharger units, exhaust manifolds, and piping. 

• Add insulation to the chiller and coolant lines at a minimum to keep the 

packages as small as possible. 

• Apply a brush-less motor driven pump for minimum package volume of 

the pump/motor assembly 

9.2.4 Compressor Efficiency 

The compressor has a major impact on system efficiency. As stated earlier, design 

modifications are recommended to improve efficiency and take advantage of R-152a 

and the secondary loop system attributes. 

9.2.5 Control Strategy 

Recently, technologies to reduce overcooling while maintaining passenger comfort 

and prevent fogging have been developed and introduced to the market. These 

technologies require extra content and added cost to the system in the form of 

externally controlled compressors and dedicated control systems. 

In contrast, when using Delphi’s Capacity Control technology with the secondary 

loop system, the benefit of reducing overcooling and consequently saving energy is 

an inherent part of the system. This Capacity Control) technology was applied to the 

secondary loop system Astra demonstration vehicle. The control capacity approach 

is inherent to the secondary loop system, only the coolant flow rate needs to be 

controlled and adjusted to achieve the desired level overcooling reduction. This 

control method can be integrated with an internally / externally controlled or fixed 

compressor. 

Because of the control capacity system demonstrated on the Astra vehicle, when 

compared to the production manual A/C system, a minimum overcooling was 

implemented and resulted in a calculated reduction in fuel use of between 13% and 
47

16% for various conditions. This estimates fuel use reduction is calculated using a 

model developed by the National Renewable Energy Laboratory (NREL) under the 

direction of John Rugh and is described earlier in section 8.3. 

9.2.6 “Cold” Storage Systems 

Another option that can be implemented in secondary loop system is the use of the 

stored cold fluid (Cold Storage) as a cold buffer to provide passenger comfort with 

the A/C compressor off. Two key transient conditions that can leverage the cold 

buffer to maintain comfort and reduced fuel use are: 

• Idle-Stop or turning the engine off under idle conditions. 

• Engine Management Systems (EMS) that can be adjusted to engage 

or disengage A/C compressor such that the energy use is reduced. In 

the case of small engine vehicles, drivability is improved with the 

compressor off. Added energy savings can be harvested with the 

compressor disengaged during accelerations and engaging it at 

maximum load to recover some of the braking energy (decelerations) 

to replenish the cold buffer. These EMS opportunities have been 

discussed with experts in engine control technology and they 

expressed that these viable opportunities need further study. 

X. SUMMARY AND CONCLUSIONS 

The goal of designing, building, and testing a commercially viable, energy-efficient, 

secondary loop HFC-152a mobile air conditioning system has been achieved. The 

system was applied to a 2007 Opel Astra to demonstrate this technology in a vehicle 

common in the European Union, wherein a phase-out of HFC-134a will begin in 

2011. 

• All component designs are based on known technologies with no anticipated 

barriers to timely production. The cooling performance of the system is equal 

or better to that of the baseline HFC-134a production system, as intended. 

The energy performance of the system is comparable to that of the baseline, 

due, in large part, to the inherent ability to easily control the cooling capacity 

of the system to provide comfort while avoiding the energy losses associated 

with overcooling, and subsequently reheating, the air. Compared to the 

production R-134a system, the energy use of the A/C compressor was shown 

through the NREL analysis to be 13% to 16% lower by applying Delphi’s 

Capacity Control algorithm 

• The Opel Astra was successfully demonstrated at the SAE Alternate 

Refrigerant Systems Symposium in July of 2007, during which attendees 

participated in an engineering review of the project and demonstration rides to 

assess the comfort of the system relative to other alternatives. 

Environmentally, by applying R-152a to a secondary loop system , Global 
48

Warming impact of direct refrigerant emissions is reduced by at least 94% 

when compared to an equivalent R-134a system. This reduction takes into 

account both the lower GWP of R-152a and the lower mass of refrigerant 

required (40% less than the production R-134a system) 

• The application of a secondary loop to vehicle air conditioning allows some 

unique fuel-saving opportunities. The secondary loop coolant system acts as 

a cold buffer that can provide cooling for a period after the compressor is 

disengaged. By combining the control of this cold buffer with the engine 

management system, the compressor can be engaged or disengaged when 

the opportunity exists for fuel savings. For example, the compressor can be 

disengaged during vehicle accelerations save energy and be fully engaged 

during vehicle decelerations to use vehicle engine braking energy to replenish 

the cooling capacity of the cold buffer. Idle-stop, i.e. turning off the engine, 

and thereby the A/C compressor when the vehicle is stooped for brief periods 

can be enhanced by the ability of the cold coolant buffer to provide comfort 

under compressor off conditions as compared to a direct system. 

Application of a secondary loop to vehicle A/C is a new concept that carries with it its 

own challenges with respect to heat exchange and system controls. Accordingly, as 

with any new system, many opportunities exist for improvements. Some of these, 

such as control technologies are highlighted in this report. 

ACKNOWLEDGMENT 

The authors would like to express their sincere thanks to Dr. Stephen O. Andersen 

and Kristen Taddonio of the U.S. Environmental Protection Agency for their support, 

as well as John Rugh of the US DOE National Renewable Energy Laboratory for his 

modeling, analysis and collaboration on the vehicle Fuel Use. Thanks also to 

General Motors for the kind assistance in supplying the vehicle for the project. The 

authors also whish to thank the Delphi Secondary Loop Team for the huge effort in 

seeing the project through. Thanks also to the Peer review team for inputs, 

comments, and support. 

REFERENCES 

[1] SNAP, “EPA-HQ-OAR-0488” Draft 

[2] M. Ghodbane, “Flammables Refrigerants”, Delphi Thermal Systems, A.T.D. 

 Group, Lockport, N.Y., April 28, 1997 

[3] M. Ghodbane, “An Investigation of Flammables Refrigerants in Mobile Air 

 Conditioning”, 1999 SAE International Congress (1999-01-08740) 
49

[4] M. Ghodbane, “On Vehicle Performance of a Secondary loop system”, 2000 

 SAE International Congress (2000-01-1270) 

[5] M. Ghodbane et Al, “2003 Alternate Refrigerant Symposium”, Scottsdale, AZ 

[6] M. Ghodbane et Al, “2007 Alternate Refrigerant Symposium”, Scottsdale, AZ 

[7] W.O. Forrest and M.S. Bhatti, “Energy Efficient Automotive Air Conditioning 

 System”, 2002 International Congress (2002-01-0229) 

[8] Improved Mobile Air Conditioning (IMAC) SAE, “Improved Efficiency” Final 

 Report, December 2006 

[9] Kays W.M., and London A. L. “Compact Heat Exchangers”, McGraw Hill Book 

Company, 2nd Edition