## 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, 20072 

### 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 Condi tions 263



### 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 484

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 Syst em 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 Sup er Heat on
Chiller Effectiveness 17 Figure 10 Typical Cooler 17 Figure 11 Cooler Heat Transfer Performance 1 8 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 O utside Air Mode (Overcooling Reduction) 27 Figure 19b Secondary Loop Capacity Control in O utside Air Mode (Overcooling Reduction) 27 Figure 20a Secondary Loop Capacity Control in O utside Air Mode, Medium Loads (Overcooling Reduction) 28 Figure 20b Secondary Loop Capacity Control in O utside Air Mode, Medium Loads (Overcooling Reduction) 29 Figure 21a Secondary Loop Capacity Control in O utside Air Mode, Low Loads (Overcooling Reduction) 30 Figure 21b Secondary Loop Capacity Control in O utside 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 Comparis on at Medium
Loads 34 Figure 24b Coefficient of Performance Compariso n at Low
Loads 34 Figure 25a Compressor Power vs Capacity at Road Load 35 Figure 25b Compressor Power vs Capacity at Idle 365

Figure 26a Speed Correction Schedule (SC03) 3 7 Figure 26b New European Driving Cycle (NEDC) 37 Figure 26c Highway Fuel Economy Test (HWET) 3 8 Figure 27 Sample Inputs to the “Advisor” Softwa re 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 Storag e 44 Figure 32 Compressor Power vs Effectiveness 4 5

List of Tables

Table 1 Test Matrix 21
Table 2 Air Flow Rates 29 Table 3 System Mass 436



### I. INTODUCTION

All currently manufactured automobile and light dut y 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 E arth’s ozone layer. While other refrigerants were studied at that time as possible replacements, HFC-134a was chosen for its combination of refrigerating propert ies and safety characteristics, despite indications that its global warming potenti al (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 th at requires the use of refrigerants with global warming potentials less th an 150 in all new-type vehicles starting in 2011 and in all new vehicles by 2017.

The chemical industry has been working to synthesiz e new chemicals and mixtures of chemicals that might one day prove useful, but h as met with limited success to date. Only two commercially available refrigerants exist today that have a GWP below the required threshold, these being carbon di oxide (R-744) and HFC-152a (R152a). Recently, DuPont and Honeywell have proposed HFO1234yf as R-134a replacement. The automotive industry will likely ha ve 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 challeng es, especially to limit refrigerant leakage. In addition, safety concerns associated with the health effects of excessive leakage into the passenger compartment wi ll likely require engineering mitigation, at least in the United States [1].

R-152a is very similar to the current R-134a in ter ms of pressure, refrigerating characteristics, and compatibility with currently u sed system components and materials, making it easier and less expensive to i mplement. It does, however, suffer from the disadvantage of being mildly flamma ble, which can be successfully mitigated by applying the refrigerant to a secondar y loop system, in which the refrigerant is contained completely within the engi ne 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]. Th is requires additional components, but is still a lower cost and lower com mercial risk option than R-744. As far as HFO1234yf is concerned, the jury is stil l out on its candidacy as R-134a replacement.

The US Environmental Protection Agency has recogniz ed the environmental opportunities R-152a can offer to the vehicle air c onditioning industry and the importance of demonstrating this technology in a ve hicle system that highlights energy efficiency and commercially viable component s. To these ends, this demonstration program was created.7



### II. BACKGROUND

Delphi has a long-standing history of secondary loo p 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 m eans of satisfying potential legislation requiring the use of low GWP refrigeran ts [3].

The first demonstration was a proof-of-concept exer cise using identical 1999 Chevrolet Malibu’s, one equipped with a secondary l oop R-134a system and the other a regular production, direct expansion R-134a system. These vehicles were demonstrated at the 1999 Phoenix Forum (currently t he SAE Alternate Refrigerant Systems Symposium) during which participants were i nvited to ride in both vehicles and see if they could identify the secondary loop v ehicle. 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 coola nt, 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 i s stored under-hood where its temperature remains near ambient, unlike the passen ger 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 Sympos ium, Delphi and Volvo jointly demonstrated a Volvo XC90 (SUV) R-152a secondary lo op system with proof-ofconcept (pre-prototype) components [5].ref. This te sting 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 Sympos ium, 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 effic ient operation.

In addition to the activities noted above, Delphi e ngineers have published numerous SAE technical papers on the topics of alternative r efrigerants and A/C system energy efficiency [7]. These are cited in the refer ence 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 me asuring 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 perfor mance and energy, and then demonstrated the vehicle with the secondary loop sy stem at the 2007 SAE Alternate Refrigerant Systems Symposium at their mid-July mee ting in Scottsdale, Arizona.



### 3.1 Project Objective

System design consists of fully understanding the c urrent physical and operational characteristics of the Astra and applying that know ledge 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 fo r 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). A dditional 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) wi ll be held to discuss initial component and system designs, as well as the intend ed test methodologies to quantitatively measure system characteristics. A se cond meeting(s) will be held during the testing phase for those interested in vi ewing the data of the secondary loop system. Outputs from these meetings that woul d contribute to improved system performance will be incorporated into the design gu ide.

Attached in appendix A: Power point file (EPA_R152a _Project_overview.ppt) containing diagrams that both reference, and furthe r describe, the new components to be designed, developed and incorporated into the system. In appendix B: The first Peer Review Team Meeting Presentation. The l ink to the presentation given at 2007 SAE Alternate Refrigerant Symposium given in S cottsdale, AZ is below: http://www.sae.org/events/aars/presentations/2007aa rsvehicles.pdf



### 3.2 Testing and Demonstration

Also included in the attached PowerPoint file is th e summary of the testing protocol for the two system architectures. These are, of co urse, 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 (LT C) 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 Collectio n 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 secon dary loop R-152a system energy efficiency will be collated into design guid elines to accompany the final report.

As requested in the statement of work, EPA will hav e all property rights and ownership of Design Manual. It is understood that p roperty rights and ownership of the Design Manual means the manual itself, as a tan gible publication, and not ownership of the intangible and potentially patenta ble ideas disclosed therein.



### IV. PRODUCTION R-134a VEHICLE AND SYSTEM DESCRIPTIO N

A 2006 Opel Astra production vehicle was used in th e 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 schematica lly in figure 2. The system is comprised of an R-134a refrigerant circuit consisti ng of components such as compressor, condenser, expansion device, and refrig erant 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) i s an excellent refrigerant. Obviously, the only negative characteristic of R-15 2a is its mild flammability. One way to overcome the safety aspect of the mildly fla mmable refrigerant is to use it in conjunction with a secondary loop cooling system. F igure 3 depicts a solid model of a Secondary loop system.

The conventional evaporator is replaced by a second ary fluid (coolant) to air heat exchanger called a cooler. The secondary loop syste m uses an R-152a refrigerant direct expansion cooling system (primary system) to cool the coolant via a refrigerant to coolant heat exchanger called a chil ler. The cold coolant is then pumped trough the cooler inside the passenger compa rtment of the vehicle to provide the necessary cooling.

In order for a Secondary loop system to be viable a lternative, the heat transfer resistance resulting from the added level of heat e xchange must be minimized through proper selection of the coolant and system design. In addition, the packaging and mass of the pump, reservoir, and chil ler must be optimized. Unlike the primary refrigerant, it is necessary for the se condary 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 in side the passenger compartment of the vehicle. There are a variety of commercially developed single-phase heat transfer fluids, which could be used as coolants. I n this investigation, ethylene glycol / water solution was chosen because it is common, w ildly used, and validated for automotive applications. The selection of the perce ntage of the ethylene glycol used in the coolant is detailed later in the report.

Figure 4 shows in a schematic fashion the lay out o f the Opel Astra Secondary loop system. The system encompasses a primary refrigeran t circuit consisting of compressor, condenser, and an expansion device coup led with the coolant circuit through the chiller. In this configuration, the chi ller is located up front, near the compressor and cools the circulating coolant. The a mount of coolant flow to the cooler inside the vehicle is regulated by a variabl e speed pump. When comparing the baseline lay out shown in figure 2 and the seco ndary loop system lay out in figure 4, the primary refrigerant suction and liqui d lines of the secondary loop system are much shorter than the baseline system. The shor t suction line introduces low refrigerant pressure drop, which translates into pe rformance 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, conde nser, and expansion device coupled with a secondary fluid circuit via a chille r. The secondary fluid circuit is composed of a reservoir, a pump, and a cooler takin g place of the evaporator inside the HVAC module. The chiller, which is thermally eq uivalent to an evaporator, is the link between the refrigerant circuit and the second ary fluid circuit.



### 6.1 Compressor

As in the direct system, the compressor still is th e heart of the secondary loop system, and consumes most of the energy. The sizin g and the selection of the compressor is critical to any system and in the cas e of secondary loop system, it is more significant because of the added thermal resis tance introduced by the secondary fluid.

If equivalent cooling capacity and similar compress or technology is used for the baseline and the secondary loop system, the displac ement of the secondary loop compressor is expected to be higher (10 to 15%) tha n the baseline compressor; however, this increase can be reduced to a single d igit if design and efficiency improvements of the compressor are realized. A theo retical evaluation of the displacement increase can be expressed by the follo wing equation:



##                        

∆ ∆        

=        

− −

− −

− − −

a R a R

a R vol a R vol

a R evap a R evap

a R suct a R suct

a R a R SL 

## RPM RPM

h h

Disp Disp

152 134

152 134

152 134

152 134

134 152 

## η η

ρ ρ14

Where D isp 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. Usin g the saturation properties of the refrigerants at 41(5), 50(10), 59 15), and 81 0 F (27 0 C) compressor in temperatures with the assumption that both compress ors have identical speed and isentropic efficiency, on the average the R-152a co mpressor displacement is higher by 11.5%.



### 6.2 Condenser

Identical condenser and fan were used for the basel ine and the secondary loop system. As in any mobile air conditioning, front ai r management is important for optimum system performance especially during engine idling situations. When the vehicle is idling, re-circulation of the warm air r aises the average air temperature entering the condenser and therefore reduces the ca pacity of the condenser. The recirculation of warm air may be also caused by insuf ficient clearance between the engine and the front-end module in which case the w arm 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 pr event 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 (co olant) heat exchanger. In a secondary loop system, the refrigerant evaporation occurs in the chiller instead of the evaporator for a baseline system. Technically t he 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 0 F (10 0 C)

Currently at the prototype stage, the chiller is a brazed aluminum plate type with enhanced heat transfer surfaces for optimum heat tr ansfer and pressure drops. Different shapes and designs are portrayed in figur e 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 0 F (10 0 C)

Figures 7 and 8 depict the thermal performance of t he 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 va riation of the refrigerant and the coolant pressure drops with the coolant flow rate a t fixed chiller coolant in temperature. As expected, the heat transfer increas es with coolant flow rate. The data has also shown that the heat transfer drops of f 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 secon dary 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 s econdary loop system is lower than the evaporator out superheat for the baseline syste m. 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 ε T =(T coolant in -T coolant out ) 

### / (T coolant in -T sat at Pout )

2.8 5.6 8.3 11.1 13.9 16.7 19.4 ( 0 C)



### 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 chi ller and is pumped to the cooler where air is cooled to the desired temperature. Fig ure 11 illustrates the thermal performance of several coolers. Again, this heat ex changer has to be compact, lightweight and erosion / corrosion resistant. Suc h heat exchanger is shown in figure 9. It is aluminum brazed flat tube and cent er 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) i s 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 pha se heat transfer fluids which

could be used as coolants. Glycols water mixtures, non-aqueous fluids such as synthetic organic fluids, Dowtherms, and Silicon oi ls are among the candidates. Because of their heat transfer degradation due thei r high viscosity at low temperatures, propylene glycol water solutions are mostly used in medium temperature applications above 0 0 F (-18 0 C). 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 applica tions 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 0 C) and 60 0 F (15.5 0 C), hence the 

### Figure 12 Coolant Pump20

evaluation of secondary fluid thermo-physical prope rties have to be taken into consideration in this temperature range. Based on h eat transfer properties, familiarity, wide automotive validation and applica tions, and the presence of inhibitors to prevent corrosion, engine coolant whi ch 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 adequ ate transport and thermophysical properties and provides sufficient heat tr ansfer while protecting the coolant circuit components from freeze damages. The EG “bur st” 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 percen t concentration if the solution is pumped through the coolant circuit. Generally au tomotive air conditioning systems are controlled not to operate when the ambient te mperature is below 40 to 45 0 F (4 to 7 0 C), 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 am bient temperatures (~ -50 0 F) when the air conditioning system is off.



### VII. PERFORMANCE COMPARISON

The function of an A/C system is to quickly and eff iciently maintain passenger’s thermal comfort irrespective of external ambient co nditions. The most severe conditions a vehicle is subjected to are the soak a nd 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 usi ng the same vehicle and tunnel under identical testing conditions.



### 7.1 Test Description and Methodology

The vehicle was first tested with its original equi pment for baseline purposes. The same vehicle was outfitted with a secondary loop sy stem with same or similar in size as the current R-134a components such as condenser.



### Table 1. Test Matrix

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

* O.S.A. = Outside Air; Rec. Air = Re-circulated (C abin 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 0 F x 40% R.H. (40 0 C) City Traffic Schedule

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

59 0 F x 70% R.H. (15 0 C) 95 0 F x 40% R.H. (37 0 C)

95 0 F x 40% R.H. (37 0 C) 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 0 F x 60% R.H. (27 0 C) 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 instrum entation for climate control tunnel testing, the vehicle was also instrumented w ith several temperature sensors and pressure transducers to monitor all the relevan t 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 identica l environment using the same
vehicular thermal tunnel. Table 1 lists the test c onditions under which both systems were evaluated. The test matrix was designed to pro vide enough performance data for comparison purposes. It is based on testing pro cedures used for validation and © 2007 Delphi Corporation, Inc.22

benchmarking of production and prototype air condit ioning systems. The first set of tests shown on the table were dedicated to high loa ds (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 circulati on inside the passenger compartment. Figure 14 shows the dynamic changes of the discharge and cabin

temperatures under ambient conditions of 115 0 F (46 0 C) x 15% relative humidity. These conditions are representative of the Phoenix area summer climate. The driving cycle consisted of a soak, followed by a ci ty 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 second ary loop system and the

Performance Comparison @ 46 0 C 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 1 4 Performance Comparison at High Loads

© 2007 Delphi Corporation, Inc.23

baseline production system are almost identical thr ough the driving cycle regardless of supplied air, being in outside air (fresh air) o r re-circulated air (cabin air). Figure 15 presents the cooling performance comparison of t he two systems at 104 0 F (40

0 C) x 40% relative humidity under soak and cool cond itions, and stable driving conditions. Again, the discharge and cabin temperat ures are almost identical through the driving cycle.



### Performance Comparison @ 40 0 

### C 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 con cern because of the thermal capacitance associated with the coolant. However, a s shown in figures 14 and 15, both systems were soaked to the same temperature, y et the response time delay (thermal lag) of the secondary loop system is negli gible. The thermal lag concerns are emanated from thoughts that the coolant tempera ture would match the soak temperature (cabin temperature) when the vehicle is soaking in the sun. However, data has shown that the coolant temperature stays a t ambient temperature since the whole system including the coolant reservoir is und er-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 ambien t conditions concluded that the thermal lag is not an issue. The thermal lag is cer tainly measurable however not



### Figure 15 Performance Comparison @ High Loads

© 2007 Delphi Corporation, Inc.24

perceptible. Some of the recent subjective evaluat ion 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 loo p system.



### 2007 SAE Phoenix Alternate Refrigerant System Sympo sium

Vent Temperature at Average Ambient: 107 0 

### F (41.7 0 

### C) 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



### ( 0 

### C) 18.3

15.5

7.2

* C.T.= City Traffic

The forum [6] was held in Scottsdale, Arizona and w as represented by almost 270 attendees. The participants came mainly from the mo bile air conditioning industry. Besides discussing the progress of alternate refrig erant technologies, the forum
covered the comfort evaluation of the demonstration vehicles. The evaluation



### Figure 16 Road Performance at High Loads25

consisted of a subjective comfort rating coupled wi th air vent and cabin temperature readings at each phase of the driving cycle.



### 2007 SAE Phoenix Alternate Refrigerant System Sympo sium
Cabin Temperature at Average Ambient: 107 0 F (41.7 0 C) x 24% R.H.

50 60 70 80 90 100

1

Driving Schedule ( 0 F)

10 21.1 37.8

( 0 C) 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 0 F (62.9 0 C)

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

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

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

* C.T.= City Traffic

The demonstration vehicles were identical (2007 Ope l 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 r ides were random and blind for the purpose of eliminating any bias for or against a ty pical refrigerant or system. Figures 16 and 17 display the objective evaluation, which i s 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 t he R-152a secondary loop system are similar in performance and are better performer s than DP-1 and AC-1 systems.



### Figure 17 Road Performance at High Loads26 2007 SAE Phoenix Alternate Refrigerant System Sympos ium

Comfort Rating at Average Ambient: 107 0 F (41.7 0 C) 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 sh own 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 Condit ions

Mobile air conditioning systems are generally desig ned and sized for maximum
loads which in most cases are soak and cool downs u nder extreme environmental
conditions. Under medium ambient conditions such as 95 0 F (35 0 C) x 40% relative humidity and mild ambient conditions such as 80 0 F (27 0 C) x 60% relative humidity, overcooling occurs in nearly all vehicles resulting in a reheating process by tempering with the temperature door, sometimes call ed blend door to maintain
comfort. This occurrence is sometimes called “serie s reheat” in the industry. In order to reduce the series reheat, today’s systems requ ire sophisticated and expensive controls to operate



### Figure 18 Subjective Comfort Evaluation at High Loa ds27 Discharge Temperature Comparison (10 min. Point)
@ 95 0 F x 40% R.H. (35 0 C) 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

( 0 C)

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

Cabin Temperature Comparison (10 min. Point)
@ 95 0 F x 40% R.H. (35 0 C) High Blower, OSA

60 62 64 66 68 70 72 74 76 78 80

1 ( 0 F)

30 mph (48 kph) 70 mph (113 kph) Idle 50 mph (80 kph) 15.5 17.8 20.0 22.2 24.4 26.7

( 0 C) 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 convention al system, operating the evaporator at higher pressures elevates the evapora tor air outlet temperature thus minimizing the reheating process and in turn reduci ng the compressor shaft power.

The secondary loop system offers this benefit of ov ercooling prevention by simply varying the coolant flow using the speed of the pum p, which in turn, reduces the compressor shaft power. This technique of “series r eheat” reduction is not complicated and is less costly.

Discharge Temperature Comparison (10 min. Point)
@ 80 0 F x 60% R.H. (27 0 C) 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

( 0 C)

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

Figures 19a and 19b depict a typical wind tunnel te st data that represents the cooling performance comparison of the baseline and the secondary loop system at
95 0 F (35 0 C) x 40% relative humidity. The tests were conducte d 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 s upply was dialed in outside air mode (fresh air). The airflow rate in outside air m ode at high blower position is
around 260 ft 3 /min. (7.36 m 3 /min.) for both systems. In re-circulated air mode at blower setting 3 which represents high medium blowe r, on the average the airflow rate for the secondary loop system is higher than t he baseline. The average amount of air flow rate in re-circulating mode for the sec ondary loop system at blower setting
3 is 216 ft 3 /min. (6.12 m 3 /min.) while the baseline air flow rate is 205 ft 3 /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 basel ine and the secondary is mainly due to the depth difference between the baseline ev aporator and the secondary loop
cooler.

Cabin Temperature Comparison (10 min. Point)

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

( 0 C) 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 discha rge 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 ba seline 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 (ble nd door) for reheating purpose. However, this maneuver consists mainly of redirecti ng 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 late r 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 variati on 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-s troking a variable displacement compressor. Figure 19a illustrates the reheating pr ocess 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 rehea ting option or in a secondary loop system with modulating coolant flow for reheating p urposes the passenger comfort is always respected and not compromised.

Discharge Temperature Comparison (10 min. Point)
@ 80 0 F x 60% R.H. (27 0 C) 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

( 0 C)

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

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

0 C) x 60% relative humidity and low ambient conditio ns, overcooling very frequently occurs and in most cases to maintain comfort, rehea ting using temperature door (blend door) levels and/or airflow reduction is nee ded.



### Figure 21a Overcooling Reduction © 2007 Delphi Corporation, Inc.31 Cabin Temperature Comparison (10 min. Point)
@ 80 0 F x 60% R.H. (27 0 C) 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

( 0 C)

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

Figures 20a through 21b show the cooling performanc e comparison of the baseline and the secondary loop at 80 0 F (27 0 C) 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 th e 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 cabi n temperatures under moderate loads portrayed by high medium blower (set ting 3) and outside air conditions. Instead of operating the pump at maximu m coolant flow, which is in this case over 4 gpm (15.14 lpm), the pump is operated a t 3 or 2.5 gpm (11.35 or 9.46 lpm). The graphs show that overcooling reduction ca n be achieved by operating the pump at 2.5 gpm (9.46 lpm) or below without comprom ising the comfort. Under lower loads where the blower is set at medium (sett ing 2) or lower, the coolant is cut down further to 2 gpm (7.57 lpm) or lower, therefor e reducing the load on the primary system and consequently the compressor shaf t work is reduced. The load on the primary system, which is mainly the compress or power, is brought down by reducing the overcooling depending on the desired c omfort. 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 0 F (40 0 C) x 40%, 95 0 F (37 0 C) x 40%, 80 0 F (27 0 C) 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 chille r for baseline and secondary loop respectively, and h is the refrigerant enthalp y 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 CO P. The outcome was

COP Comparison @ 40 0 C 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 in 

### (Heat Exchanger) 

# h out 

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

1 1.2 1.4 1.6 1.8 2 2.2 2.4

1 COP

30 mph (48 kph) 70 mph (113 kph) Idle 50 mph (80 kph) 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 t he coefficient of performance (COP) is the ratio of cooling capacity versus the e nergy 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 purchase d) 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 e conomically 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 p erformances (COPs) of the baseline system with several control capacity optio ns of the secondary loop system at medium and low loads. The graphs clearly show th at 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 eff iciency of mobile air conditioning (IMAC) [8]. The lower COPs for the sec ondary loop system are 

### Figure 23 Coefficient of Performance (COP) Compari son © 2007 Delphi Corporation, Inc.34

attributed to two factors, the lower compressor eff iciency at reduced compressor strokes and compressor operation with R-152a instea d of R-134a.

Coefficient of Performance (COP) Comparison (10 mi n. Point)
@ 80 0 F x 60% R.H. (27 0 C) 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 mi n. Point)
@ 80 0 F x 60% R.H. (27 0 C) Ambient & Blower Setting 2

0 0.5 1 1.5 2 2.5 3 3.5 4

1 C O P

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) Compar ison

Figure 24b Coefficient of Performance (COP) Compar ison © 2007 Delphi Corporation, Inc. © 2007 Delphi Corporation, Inc.35

As in IMAC study, the energy usage or compressor po wer is a better metric than COP to evaluate fuel usage and emissions. The cooli ng capacity is directly proportional to the cooler (evaporator) air out tem perature. The capacity reduction is accomplished by increasing the cooler (evaporator) air out temperature while keeping the comfort at the desired level. For the s econdary 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 Mod e

Ambient: 80 0 C (27 0 C) 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 OutCompressor 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

( 0 C)

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 v ariation with the cooler (evaporator) air out temperature. The data points o n the left and the right of the graphs illustrate the maximum cooling capacity (low est 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 mentione d 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 overco oling. The capacity is controlled to match the desired comfort without relying on reh eating. As shown in figure 25a at road load, if the cooler air out temperature is con trolled to 46 0 F (8 0 C) 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 0 F (9.5 0 C). As seen on the plots, the secondary loop system compressor power c an be maintained at equal or lower values of the baseline compressor power by me ans 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 12Compressor 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

( 0 C)

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 0 C (27 0 C) 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 achiev ed through modulating the coolant flow rate. At medium to low loads, the seco ndary loop system capacity control is an effective way of reducing the compres sor shaft power. The next section addresses the benefit in fuel consumption g enerated 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 ai r conditioning fuel use of a baseline R-134a and a secondary loop R-152a for various driv ing 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” Softwa re39

The inputs to the model consisted of the compressor power versus speed generated
from vehicle tunnel tests at ambient of 80 0 F (27 0 C) x 60% relative humidity, the vehicle characteristics such as mass and engine typ e, and the drive cycle data. The driving schedules simulated in this analysis are th e New European Driving Cycle (NEDC), the Speed Correction Cycle (SC03), and High way 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, m echanical accessory load, and rolling resistance coefficient and engine character istics. 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 of f 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 o f 80 0 F (27 0 C) x 60% relative humidity, vent mode, full cold and the blower setti ng 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 c oolant 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 t he R-152a secondary loop system with control capacity on. As expected the total fue l consumption with A/C off at each driving cycle is the lowest, however with A/C on, t he consumption is less with the control capacity secondary loop system than the bas eline 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 loo p system with controlled capacity. Depending on the driving cycle, the advan tage of the secondary loop with controlled capacity ranges from 13 to 16%. This fue l saving can be achieved by reducing overcooling and eliminating the unnecessar y reheating process to reach comfort. As stated before, averting the overcoolin g would ultimately translate into saving energy, hence less fuel use and less emissio n.

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 HEFETFuel 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 on41



### IX. SECONDARY LOOP SYSTEM DESIGN GUIDELINES

For comparison purpose, it is recommended to baseli ne the vehicle before equipping it with Secondary Loop system.



### 9.1 Overall System

• To meet maximum performance requirement, it is rec ommended to use higher displacement (10 – 15%)compressor

Ex: Assuming same speed & isentropic efficiency for both compressors and using saturation properti es of the refrigerants, @ 5, 10, 15 & 27 0 C compressor in temperatures, the increase in displacemen t 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 HEFETA/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 HEFETA/C Fuel Use (L/100 km) R134a

SL R152a w/ CC

-13% -13%

-16%



### Figure 30 A/C Fuel Consumption42 (recalibrated high side pressure transducer for R- 152a. It is usually located around condenser)

• Check front end air flow (minimize air re-circula tion)

• The TXV needs to be reset to match R-152a propert ies and the secondary loop applications

• Cooler coolant in temperature (instead of cooler a ir off) is
recommended for freeze control

• Capacity controller

• Insulate refrigerant in line to chiller

• Insulate chiller

• Depending on the length, coolant lines could be in sulated

• 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 chi ller)
 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 guidel ines are described in the design improvements section.



### 9.2 Potential Design Improvements

As with all design efforts, optimization is a serie s 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 m inimum 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 Savings43

Each of these features will be summarized in order to give options for performance and/or a reduction in the energy needed for operati on.

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 demonst ration vehicle content
Cooler Similar to evaporator Based demonstration ve hicle 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 ma jority of the added mass of the system. As an improvement, a reduction in mass ca n be achieved by optimizing the designs. Mass of the chiller depends on several fa ctors: its required size to meet the heat transfer requirements in the vehicle, the nece ssary material gage to meet field reliability targets, its mounting features in the v ehicle, its plumbing to interface with primary refrigerant and the coolant circuits. The a mount 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 chille r design based on these
characteristics are:

1. Select the chiller size to meet the cooling requ irements, as well as,
balance the heat transfer and the chiller effective ness 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 c oefficient, and the refrigerant / coolant flow rates.

2. The gage of the material selected must be consis tent with the requirement of the SAE J639 specification, the mate rial gage must meet the manufacturing, and durability needs as spe cified by each
OEM.

3. Chiller mounting features such as brackets and f ittings have an impact on the mass of the chiller. Directionally, it is pr eferred 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 prim ary and secondary fluids will depend on the type of refrigerant contr ol 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 a nd the quantity of
insulation.

The major source of additional mass is the amount o f 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. A dditional coolant can be added to provide extra “cold” storage. For the Astra vehi cle, 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 syste m, the Astra system was able to maintain vehicle comfort up to three times longer t hen the baseline R-134a production system. This level of performance was me asured 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 benefit s as we have discussed, the
coolant circuit adds a “heat transfer resistance” t hat must be overcome to provide the necessary temperature difference. One common me asure of how well heat is transferred in a given heat exchanger is its effect iveness. 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 sy stem and the effectiveness of the heat exchangers at a fixed system load and rang e 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 comb ined effectiveness of the cooler/chiller as compared to the direct expansion system evaporator effectiveness reduces the power required. The ratio for equivalen t 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 Evaporat or Effectiveness Ratio Assumes: Same Compressor Speed, Isentropic Efficien cy, 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 EvapComp Power SL/Comp Power Base

Road Load Condensing T Idle Condensing T Evap T3 (W R152a-SL /W R134a )=( א h evap-R134a * א h isen-R152a * א isen-R134a )/( א h evap-R152a * א h isen-R134a * א isen-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 Evaporat or Effectiveness Ratio Assumes: Same Compressor Speed, Isentropic Efficien cy, 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 EvapComp Power SL/Comp Power Base

Road Load Condensing T Idle Condensing T Evap T3 (W R152a-SL /W R134a )=( א h evap-R134a * א h isen-R152a * א isen-R134a )/( א h evap-R152a * א h isen-R134a * א isen-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 hea t exchanger effectiveness will result in lower energy requirements for the seconda ry loop system.

9.2.3 Component Package Size

Packaging the components that make up a secondary l oop system is one of the challenges in applying this technology to productio n vehicle applications. An advantage of the secondary loop system components i s that they do not require a specific in-vehicle location to perform the require d 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 sour ces 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 p ackage volume of the pump/motor assembly

9.2.4 Compressor Efficiency

The compressor has a major impact on system efficie ncy. 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 introdu ced to the market. These technologies require extra content and added cost t o the system in the form of externally controlled compressors and dedicated con trol systems.

In contrast, when using Delphi’s Capacity Control t echnology with the secondary loop system, the benefit of reducing overcooling an d consequently saving energy is an inherent part of the system. This Capacity Contr ol) 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 leve l 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 min imum overcooling was implemented and resulted in a calculated reduction in fuel use of between 13% and47

16% for various conditions. This estimates fuel us e reduction is calculated using a model developed by the National Renewable Energy La boratory (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 t o provide passenger comfort with the A/C compressor off. Two key transient conditio ns that can leverage the cold
buffer to maintain comfort and reduced fuel use are :

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

• Engine Management Systems (EMS) that can be adjust ed to engage or disengage A/C compressor such that the energy us e is reduced. In the case of small engine vehicles, drivability is i mproved with the compressor off. Added energy savings can be harves ted with the compressor disengaged during accelerations and enga ging it at maximum load to recover some of the braking energy (decelerations) to replenish the cold buffer. These EMS opportuniti es have been discussed with experts in engine control technology and they expressed that these viable opportunities need furt her study.



### X. SUMMARY AND CONCLUSIONS

The goal of designing, building, and testing a comm ercially viable, energy-efficient, secondary loop HFC-152a mobile air conditioning sys tem has been achieved. The system was applied to a 2007 Opel Astra to demonstr ate this technology in a vehicle common in the European Union, wherein a phase-out o f HFC-134a will begin in
2011.

• All component designs are based on known technolog ies with no anticipated barriers to timely production. The cooling perform ance of the system is equal or better to that of the baseline HFC-134a producti on 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 easi ly control the cooling capacity of the system to provide comfort while avoiding the energy losses associated with overcooling, and subsequently reheating, the a ir. 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 th e SAE Alternate Refrigerant Systems Symposium in July of 2007, duri ng which attendees participated in an engineering review of the projec t and demonstration rides to assess the comfort of the system relative to other alternatives. Environmentally, by applying R-152a to a secondary loop system , Global48 Warming impact of direct refrigerant emissions is r educed 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 syste m)

• The application of a secondary loop to vehicle air conditioning allows some unique fuel-saving opportunities. The secondary lo op 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 examp le, 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 di rect 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 sy stem controls. Accordingly, as with any new system, many opportunities exist for i mprovements. Some of these, such as control technologies are highlighted in thi s report.



### ACKNOWLEDGMENT

The authors would like to express their sincere tha nks to Dr. Stephen O. Andersen
and Kristen Taddonio of the U.S. Environmental Prot ection Agency for their support, as well as John Rugh of the US DOE National Renewab le 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 Lo op 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”, Delph i 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 (19 99-01-08740)49

[4] M. Ghodbane, “On Vehicle Performance of a Se condary loop system”, 2000 SAE International Congress (2000-01-1270) [5] M. Ghodbane et Al, “2003 Alternate Refrigera nt Symposium”, Scottsdale, AZ

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

[7] W.O. Forrest and M.S. Bhatti, “Energy Effic ient 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, 2 nd Edition