Document ID: EPA-HQ-OAR-2009-0472-0156
Agency: epa
Document Type: Supporting & Related Material
Title: 
Posted Date: 2009-09-28T04:00Z

I. Executive Summary

Project Objective:

This project is an investigation of efficiency of R134a refrigerant
systems for mobile air conditioning.  The main objective for Team 2 is
to evaluate the COP of R134a refrigerant systems with components and
controls optimized for efficiency at moderate and low loads but with at
least equal high load performance to the Phase II Alternate Refrigerant
Cooperative Research Program [ARCRP] systems.  The goal of Team 2 was to
demonstrate an improvement of 30% in system COP over the R134a System
used in Phase II ARCRP.  In the opinion of the OEM members of Team 2,
this system represents efficiency significantly above the typical system
used in production vehicles in 2005.  Please note the discussion on pg 6
"Energy vs COP"; energy is a better metric because COP does not account
for the change in magnitude of energy usage with varying climatic
conditions.

Project Overview:  Purpose and Background:

The Improved Mobile Air Conditioning (IMAC) Cooperative Research Program
(CRP) began in August, 2004.  The efficiency team's technical approach
started with “brain-storming” ideas for potential technologies that
would improve efficiency.  This list was divided among the members and
each member was asked to research individual technologies and assesses
their potential benefits.  This research was then discussed in the team
meetings and a consensus was reached on the potential percentage
improvement that each technology offered.  Based on this assessment, the
items were prioritized, and potential suppliers of these technologies
were requested to submit sample components.  This report presents the
results of the evaluations that have been conducted on a System Test
Stand in the lab and a demonstration vehicle.

The project began with organizational meetings in the Fall of 2004.  The
team then met regularly during 2005 and 2006 to review the progress of
testing of the various technologies.  An interim report was provided to
the sponsors in December, 2005 and an update of this was presented in
June, 2006.  Vehicle development began in Spring, 2005 and continued
through the end of 2006.

The report is organized as follows:

Executive Summary

System Test Bench Evaluations 

Vehicle Evaluations

Appendices:

IMAC2-A	System setup tests

IMAC2-B	B-o-B detail data

IMAC2-C	Table of variables for bench testing

IMAC2-D	Table of variables for the vehicle

IMAC2-E	Vehicle data table

IMAC2-F	Calculations for the bench test

IMAC2-G	Vehicle COP and Capacity Calculations

IMAC2-H	Details of algorithm for vehicle

IMAC2-I	J2765 Procedure for measuring system COP [Coefficient of
Performance] of a mobile air conditioning system on a test bench

IMAC2-J	J2766 Standard for Annualized Climate Calculation of System
Power Consumption of a mobile air conditioning system

IMAC2-K	User’s manual for the vehicle controls

Summary of Results and Conclusions:

The reference baseline changed during testing and the technologies were
not always additive in final Best of Best [B-o-B] system result.  The
team decided after analyzing the impact of COP as compared to energy
consumption that it would be better to use compressor energy consumption
as a measure of success.  Comparing energy, the IMAC final best of best
system run at 10°C evaporator out temperature demonstrated over 36%
reduction in energy consumption as compared to the ARCRP baseline which
was run at lower evaporator out temperatures.  The COP improvement was
12.6% when compared at this same condition.  At equal capacity, the COP
improvement was 15%.  [The improvement is greater because the compressor
efficiency is improved at higher displacement.]  This is caused by the
reduced efficiency of the variable compressor at reduced strokes.  The
raised evaporator temperature represents a reduced reheat strategy.

ARCRP Base	IMAC B-o-B

Teao= ARCRP	IMAC B-o-B

Teao= 10C

Compr Energy	COP weighted	Compr Energy	COP weighted	Compr Energy	COP
weighted

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

-12.3%	15.0%	-36.4%	12.6%

At various intermediate steps, the efficiency improvement was measured
against a baseline system that was assembled at the beginning of the
IMAC project.  This baseline was comprised of components from the
previous Alternate Refrigerant Cooperative Research Program, but
included a new condenser.  The new condenser was configured by a new
supplier to explore different subcooling pass configurations.  The
deliverables as measured against this original IMAC baseline system
nearly met the program goals as shown in the table below.  COP was much
improved, and energy consumption nearly met the 30% target.  The
original IMAC system was not as efficient as the ARCRP Phase II system. 
These results are from the technology screening test matrix that is less
comprehensive than the full matrix used for the previous table.

IMAC Base Compr energy	IMAC B-o-B Compr energy	IMAC Base COP	IMAC B-o-B
COP

663,495	481,791	2.20	3.39

	27.4%

54.3%

Individual component technology changes showed even more improvement
relative to the baseline IMAC system.  

Standards have been delivered that document the procedures for these
measurements and annualization.  A cost/benefit analysis has been
completed to compare the technologies.  

Performance of this enhanced system has been demonstrated on a vehicle
to match the current production system.  An algorithm was developed and
demonstrated on this vehicle that produced a real time calculation of
system COP that could be used to optimize system controls for COP or
capacity.  The results of this algorithm correlated well with the bench
test results.

Outside the original scope of the project, methods for measuring a fixed
compressor with cycling clutch controls to a variable compressor have
been developed.  This demonstrated that variable compressors are more
efficient than fixed compressors.  Some new methods were developed to
accurately measure the capacity of systems under cycling conditions.  

Below is a list of the chosen technologies and their relative values for
both energy reduction, COP improvement, and a relative cost ranking as
developed by IMAC Team 2.  Relative costs are the judgment of the Team
members and are not based on any actual cost information.  Energy
improvement for some technologies can vary significantly depending on
the baseline components (e.g. results for superheat control and internal
heat exchangers can vary widely with different evaporators).



IMAC System Test Stand Results

Description of Technology	Annual COP Improvement [%]	Annual Energy
Improvement [%]	Cost Ranking Low=1, High=10	COP/Cost	Energy/Cost
Comments

Improved Condensers	6.3	-7.4	4.0	1.6	-1.9	 

	 	 	 	 	 	 

Improved Evaporators	21.2	-14.1	4.0	5.3	-3.5	 

	 	 	 	 	 	 

Reduced Reheat [with External Controlled Variable Compressor]	-2.4	-24.1
6.7	-0.4	-3.6	Saves significant energy with a small decrease in COP –
shows that energy is the correct metric.

	 	 	 	 	 	 

Internal Heat Exchanger	21.4	-12.3	6.5	3.3	-1.9	Results depend on
evaporator design.

	 	 	 	 	 	 

Cycling Fixed Compressors versus Variable Compressor 	 	 	 	 	 
Special study at request of Core Team – not considered for Best of
Best 

Fixed Piston	-18.7	+27.2	NA	 	 	 

Fixed Scroll	-11.1	+19.5	NA	 	 	 

Insulated Suction Line	1.0	-1.0	2.7	0.4	-0.4	Improvement was considered
minimal- only two points run

	 	 	 	 	 	 

Improved variable compressors, some with integrated oil separator	39.8
-18.3	6.7	5.9	-3.7	Limited test matrix 

	 	 	 	 	 	 

Condenser Subcooling Pass Design	6.3	-7.4	2.5	2.5	-3.0	 

	 	 	 	 	 	 

Subcooling Control Algorithm	1.0	-1.0	7.5	0.1	-0.1	Improvement was
considered minimal- only two points run

	 	 	 	 	 	 

Superheat Control Algorithm [enabled by use of EXV]	28.6	-16.5	7.2	4.0
-2.3	Change from 7K to 1K at low loads.  Cross charge TXV is similar. 
Results depend on evaporator design.

	 	 	 	 	 	 

Best of Best System	 	 	 	 	 	 

versus ARCRPII	15.0	-36.4	 	 	 	 

Vehicle Results:

The main objective of the vehicle was to demonstrate that the improved
control with optimized efficient components could be implemented into a
production vehicle without degrading the maximum system performance. 
The graph below, taken from the final report of the 2006 Phoenix
Alternate Refrigerant System Symposium, illustrates that the IMAC
Optimized refrigerant vehicle did demonstrate equal capacity in the
vehicle ride program. 

Systems Technologies: 

Except for the evaluation and comparison of the fixed compressor
designs, all systems used a variable displacement compressor and a
manually controlled electric actuated stepper motor expansion valves
[EXV].  The systems evaluated here were composed of the following:

Compressor

IMAC Initial-Externally variable compressor

Alternate Refrigerant CRP Phase II-same

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

Evaporator

IMAC Initial-Micro-channel design

Alternate Refrigerant CRP Phase II-Enhanced Plate and Fin design

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

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

IMAC Initial-Micro-channel design

Alternate Refrigerant CRP Phase II-Micro-channel design

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

The reference baseline system for individual technologies was not always
constant.  The test matrix used to evaluate these individual
technologies was a reduced matrix of the full IMAC test matrix and when
the Annualization was conducted the results were not comparable to
results of the full matrix.  But, the relative rankings within each
technology assessment should be the same.  These technology assessments
are discussed in more detail in section II.

Energy vs COP:

During the project development, it was recognized that energy is a
better metric than COP based on life cycle analysis.  Weighting COP gave
too high a value to low load conditions where energy consumption is low.
 Energy consumption weighting better assesses the impact on the vehicle.

The steps to calculate include the following:

Calculate average capacity of all tests of a particular technology

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

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

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

Results presented above show both the COP and energy consumption
annualized results.  Ambient weighting is based on Visteon ambient
analysis and assumes the AC is on above 10°C.  The compressor RPM
weighting is based on the weighting from a typical vehicle running on
the combined FTP City and Highway driving cycles.

An example is shown below of the difference that was observed between
two improved compressors [#4/#5] as compared to a baseline [#1],
illustrating how COP and energy was weighted for COP and energy at each
test point and then an average overall weighting.

System Test Stand-Test Descriptions:

Tests included evaluation of cooling capacity and efficiency [COP] of
each of the systems at four different speeds [900, 1500, 2500, 4000 RPM]
and seven different load conditions.  Comparisons were conducted at
stabilized operation and equal capacity at the reduced load points as
measured by the average evaporator out nozzle temperature.  System
capacity was targeted to maintain a 10(C air outlet at these reduced
load conditions.  At the high load points, the compressor was set to
maximum capacity.  

Test conditions were normalized in several aspects following the
procedures established in the SAE ARCRP Phase II program prior to
testing under stabilized conditions. Below is a summary of those areas
evaluated:

Verify air velocity and temperature profiles at the inlet to each heat
exchanger, and heat exchanger airside pressure drop.

Oil circulation was measured and adjusted per the compressor supplier
recommendation prior to testing of the final configuration of
components.

After the initial check, a system charge was selected per the charge
determination procedure as outlined in the ARCRP Phase II final report. 
Details will be presented later in this report.

Charge determination

Standards:

The Efficiency Team developed a standard for the laboratory test method
for testing mobile air conditioning systems for system efficiency.  This
team also developed an additional standard to standardize the method for
Annualization of this energy consumption.  Drafts of the two standards
can be found in the appendix.  They are as follows:

J2765 Procedure for measuring system COP [Coefficient of Performance] of
a mobile air conditioning system on a test bench

J2766 Standard for Annualized Climate Calculation of System Power
Consumption of a mobile air conditioning system

Cost vs Benefit summary:

The team has decided that the benefit part of the equation will be based
on an annualized number for energy efficiency.  The method used to
determine this has been written into the SAE standard.

The cost portion of the equation was a consensus ranking of each of the
technologies on a 1-10 scale as it is not possible to readily assess the
specific implementation costs for each technology but the team feels we
can assess the relative costs.  The results are reported in the table
above.

Acknowledgements:

Team two would first like to acknowledge the support of the sponsoring
companies for funding of this Cooperative Research Program.

We would like to acknowledge the support of the University of Illinois
ACRC (Air Conditioning and Refrigeration Center) team members who
completed the testing of the various system technologies during this
program.

U of I Team Members

Pega Hrnjak, Professor 

Xinzhong Li, Research Scientist

Stefan Elbel, Graduate Research Assistant

Jing Che, Visiting Scholar

Radko Brock, Visiting Scholar

Jacob Virostko, Undergraduate Research Assistant

Team two would also like to acknowledge the donation of a Cadillac STS
vehicle from General Motors for demonstration of the improved efficiency
system and control algorithm.

Finally we would like to acknowledge all of the companies that
contributed components for evaluation during the testing of these four
alternative refrigerant systems.  Due to confidentiality concerns, their
names will not be disclosed here. There were some components that were
submitted and not used based on preliminary evaluations.

Team Two Individual Members

Name	Company

Jacob Bayyouk 	Sanden

David Barwin	Nissan

Mahmoud Ghodbane	Delphi

Hans Hammer	Audi

John Meyer	Visteon

Mark Spatz	Honeywell

Steve Lepper	Ford

Gene Dianetti	Parker

Pankaj  Mahajan	DCX

Kwangtaek Hong	DCX

Markus Wawzyniak	Behr

Bill Hill	GM

Dennis Littwin	Fujikoki

Jan Zeng	Denso

II. System Test Bench Evaluations

Technologies Considered:

Team 2 began by brainstorming a list of technologies for consideration. 
Technologies were assigned to team members to research.  Team meetings,
in 2004, reviewed the potential benefits and ranked the technologies for
COP improvement.  This process resulted in the selection of the most
promising technologies for testing.  Below is the final evaluation
matrix which lists the technologies considered and their assessment. 
Appendix A of the Dec, 2005 Interim Report includes the detailed
assessments.

1.  Improved Heat Exchanger Designs (Condensers and Evaporators)

Presentations were given by Behr and Delphi on the opportunities for
improvement. Based on these presentations, the team decided to pursue
this technology for investigation.  All suppliers were then requested to
supply data on improved designs.  This data was used by the OEM members
to select heat exchanger designs for test.  The specification allowed
the option for thicker condensers and evaporators provided air side
pressure drop was the same as baseline.  

2.  External Control of Variable Compressors

The externally controlled variable compressor enables the system to
raise the evaporator out pressure and correspondingly, the air discharge
temperature to reduce the energy consumption of the compressor at low
loads.  The benefit for external control of evaporator outlet air
temperature was not tested as a separate item.  The individual benefit
was ranked based on the GM presentation, below is a summary:

ARCRP data suggests a 10-20% COP improvement for external controlled
compressor as compared to internal control.

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

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

The team agreed that 15% was a conservative estimate supported by ARCRP
data.  The Best of the Best I-MAC System was tested at elevated
evaporator air out control temperatures and at the normal control
temperatures used in ARCRPII.

3.  Internal Heat Exchanger

This technology was presented by Visteon and supported by the University
of Illinois.  Initially, a commercial design was tested in two
configurations to substantiate potential improvement.  Subsequently, two
automotive designs were tested.  These tests included studies of the
impact of various superheat control strategies.

4.  Scroll Compressor

Initially, the team decided not to pursue this technology based on
discussion between a key supplier and the OEM’s.  Later, in a Core
Team meeting it was proposed that Team Two should take on the evaluation
of fixed vs variable compressors.  With this assignment, team two
decided to pursue a comparison of fixed scroll vs fixed piston
compressors as compared to a variable compressor baseline.  

5.  Fixed vs Variable Compressor

A summary of the available public information was researched and
presented by Ford showing results from no difference to significant
advantage for internal control variable.  While this study was
considered beyond the initial scope of I-MAC, it remains a subject of
great interest given potential regulations and the lack of unanimous
agreement.  The Core team then requested that team two perform some
evaluation of these technologies.  Both fixed piston and scroll designs
were tested.  These tests required the implementation of new high speed
data acquisition.

6.  Oil Separator

This technology is included in some of the optimized compressor
efficiency designs and is included in item 12 below.  It was not
considered as a stand alone technology.

7.  Ejector System

Ejector technology was presented by University of Illinois personnel. 
While there is potential benefit, the technology was rated to be a low
priority due to higher complexity compared to other options.

8.  Airflow Imbalance at condenser air inlet

This item was tabled as its benefits can be applied on all refrigerant
systems.  Airflow imbalance was evaluated in Phase I of the ARCRP and
the impact on efficiency was not significant.

9.  Increased use of Interior re-circulated air

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

10.  Reduced Front End Re-circulation

This item was tabled as its benefits can be applied on all refrigerant
systems.  The impact was studied with the final system configuration by
running at elevated air inlet temperatures at the idle points to compare
back to the performance of Phase II ARCRP.

11.  Insulated Suction Line

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

12.  Optimized Compressor Efficiency

Compressor manufacturers were asked to provide prototype high efficiency
externally controlled variable compressor designs.  Two compressor
suppliers provided both a 1st generation and a 2nd generation of
improved compressors for evaluation.  Some designs include integrated
oil separators.

13.  Optimized System Control Strategy

GM presented ideas for controlling sub-cooling at lower evaporator loads
(discussed in the results section in more detail).   Varying sub-cooling
with load was believed to offer meaningful COP improvement.  The size of
the condenser sub-cooling region was believed to be of importance.  In
preliminary testing, it was discovered that control of evaporator out
superheat also had a large impact on system efficiency at low loads. 
All of these items were included for testing and are discussed
separately in the results section below.

14.  Reduced Suction Line pressure drop

Parker Hannifin presented some lab test data on this item.  The impact
can be very dramatic.  The baseline system was judged to already be on
the low side of normal.  Therefore, this item was not included for test.

15.  Flash gas removal

University of Illinois personnel indicated some potential benefit. 
While this technology is interesting, it seems to be more complicated
and the team decided not to include in this evaluation.

16.  Digital scroll

Nissan presented data on this subject.  This technology applies only to
electric driven compressors, and for this reason, this item was not
included for test.

17.  Improved evaporator air temperature distribution

All testing was done with uniform air temperature inlet to evaporator
and condenser.  The team consensus was this issue applies to all systems
and there was not a separate evaluation of this effect during this
research program. 

18.  Matching components to compressor

The team decided that since matching system components is standard
industry practice this item should not be part of the test program.

19.  Electronic Expansion Valve

This technology is an enabler to provide control of evaporator out
superheat.  It was not tested as a separate item.

20.  Electric Compressor

Since there are no (known) plans for high voltage on main stream
vehicles, the team decided not to pursue testing.

21.  Expander

This subject was researched by Honeywell and information was also
received from Refrigeration Development Company Ltd.  This technology
was not deemed to provide sufficient improvement to warrant the added
complexity.

Technologies Tested:

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

Description of Technology

Improved Condensers

Improved Evaporators

External Controlled Compressor [Reduced Reheat]

Internal Heat Exchanger

Cycling Fixed Compressors versus Variable Compressor - NOT for BoB
[Piston and Scroll type compressors]

Insulated Suction Line

Improved variable compressors, integrated oil separator

Condenser Subcooling Pass Design

Subcooling Control Algorithm

Superheat Control Algorithm w/EXV

Best of Best System

The improvement results for individual technologies are compared to a
baseline system.  Due to testing progression over a two year period,
baselines for individual technologies are not the same.  In some cases,
the baseline changes within a technology test.

The Best of the Best (BoB) system includes:

BoB Improved Condenser

BoB Improved Evaporator

External Controlled Compressor (elevated evaporator air off
temperatures)

BoB Improved Variable Compressor

Superheat Control Algorithm w/ EXV

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

The Fixed versus Variable compressor test was a special study at the
request of the Core Team.  Fixed compressors were not considered for the
BoB system.  As such, the fixed results will be presented separately
after the other results.

Test Facility:

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

The experimental facility, shown below, has two environmental chambers
and a compressor chamber in-between.  Each of the two chambers contains
a wind tunnel with variable speed blower and temperature controller,
enabling a wide range of airflow rate and air temperature for the
condenser and evaporator in the chambers.  The evaporator chamber also
has a steam supply and humidity controller to provide the latent load. 
Two chambers were designed to give three independent methods to
determine system capacity: refrigerant side, airside, and calorimetric
chamber balance.  With the three methods used, instead of two as
required by all applicable standards, the determination of system
capacities is more reliable.  In some cases, when the evaporator
refrigerant exit is two-phase, or the system operates unstably (under
some low load condition), only the airside balance is available. Careful
design and calibration were made so that at least two of the three
independent procedures could provide an agreement of ±3% as opposed to
widely accepted standard of ±5%.  

The compressor chamber holds the compressor at a desired temperature to
simulate temperature conditions in the vehicle engine compartment under
the hood. Compressor power is obtained by shaft torque and speed
measurement. A torque meter is located between the compressor and the
driven clutch, eliminating belt and clutch losses. 

 

Schematic of the experimental facility

Instrumentation schematic

Methods of Capacity and COP Measurement:

Evaporator Capacity is measured using three independent methods:

Air Capacity = dry air mass flow rate * air enthalpy difference –
condensate enthalpy (small)

Refrigerant Capacity = refrigerant mass flow rate * refrigerant enthalpy
difference

Chamber Capacity = electric heat – glycol cooling (required at low
loads) + steam supplied energy – evaporator/cooling core condensate
energy  +/- heat gain/loss through the chamber walls (small)

 

Capacity Measurement Schematic

Reported Capacity, used to calculate COP, is the average of Air Side and
Chamber Capacity EXCEPT where noted in the results.  At low loads, Air
Side Capacity is used because it is the only reliable method.

Compressor Power is calculated from measurement of shaft torque and
speed.  A strain gage torque meter is mounted in line with the
compressor through a flexible coupling.  Speed (rpm) is measured with an
optical pulse counter. 

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

Power for the condenser/gas cooler fan and evaporator blower was not
included in the COP calculations.  Heat exchanger dry air pressure drops
were close to each other.  Therefore, power for the fan and blower would
be similar between the systems.

Testing fixed compressors during cycling required high speed data (10
Hz) acquisition.  Power measurement was a straight forward high speed
data acquisition and integration of torque and rpm.  Capacity
measurement, however, presented an issue.  During cycling, refrigerant
enthalpy is not accurate.  As noted above, at the lowest loads only Air
Side Capacity is reliable.  For cycling conditions, dry bulb and
humidity leaving the evaporator were integrated separately from the high
speed data to calculate sensible and latent capacities.  The issue
encountered was that the humidity measurement used a chilled mirror
which is a slow response device itself and the air sample must travel
some distance into the device.  The chilled mirror is highly accurate at
the high relative humidity leaving the evaporator.  A review of faster
response devices revealed that none are sufficiently accurate at these
high humidity conditions.  For the conditions where Chamber Capacity is
available, a comparison was made with Air Side Capacity measure as just
described.  After the system stabilized to repetitive cycles, data was
taken and integrated over at least 10 whole cycles.  Accuracy assessment
is below.

Overall, the assessment of results indicated the following:

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

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

Compressor power accurate to ±20W.

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

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

Selected error analysis of Capacity and COP accuracy are shown in the
graphs below for both fixed and variable compressor.  

Details of Test Procedures for Capacity and Energy: 

All tests for system COP and component performance were done at steady
state (stable) conditions. The SAE ARCRP Phase II test matrix was
reduced to more efficiently screen various technologies under
consideration. The intention is to run complete SAE ARCRP Phase II test
matrix with the combination of technologies that provides the optimal
system efficiency. The test matrix is shown in Table III-2 below. These
steady state conditions cover a significant range of typical vehicle air
conditioning operation. 

Systems were compared at equal capacity by controlling evaporator air
outlet temperature as measured at the airflow nozzles at reduced load
points. System capacity was controlled by varying compressor
displacement.  The evaporator air outlet temperature was controlled to
10ºC.  The compressor was run at full capacity at 35°C and 45°C test
ambient conditions.  

Compressor chamber temperature set to 30K above condenser air inlet
temperature.

Controls Discussion:

For variable compressor operation:

The compressor was controlled with a PWM [Pulse Width Modulated]
control.  This was used to allow control of the displacement so points
with equal evaporator out air temperature could be run.  The
displacement control was based on air out of evaporator [actually the
average of the two thermocouples located in the nozzle downstream of the
evaporator].  The expansion valve controlled superheat in the range of
6-8K at high load conditions and 2-4K at low load conditions.  

For fixed compressor operation:

For the fixed compressor system, besides the change of compressor, the
EXV was replaced by a TXV as it is difficult to control Evaporator
superheat during cycling with a manual EXV. The maximum cycle rate was
specified as 6 cycle per minute for the fixed compressors running at
cycling mode. The evaporator out air temperature (measure at nozzle) was
used as the input signal for a PID controller to control compressor
clutch. The temperature variation band was specified as ±1(C. However,
this band was allowed to be widened up to ±3(C (or other number) to
keep the cycle rate lower than 6 cycles/minute. 

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

Test Name	Ambient temp. [oC/oF]	Compr speed [rpm]	Condenser	Evaporator

	Inlet air temp.      [oC/ oF]	Flow rate [l/s/CFM]	Inlet air temp.     
[oC/ oF]	Flow rate [l/s/CFM]	Simulated air selection	RH [%]	Dew point
[oC]

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

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

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

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

L45	“	1500	“	689/1460	“	130/275	“	“	“

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

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

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

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

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

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

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

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

L35b	“	1500	“	689/1460	“	60/127	“	“	“

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

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

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

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

L25a	“	1500	“	689/1460	“	60/127	“	“	“

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

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

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

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

L25b	“	1500	“	689/1460	“	60/127	“	“	“

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

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

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

L25c	“	1500	“	689/1460	“	40/85	“	“	“

M25c	“	2500	“	1067/2260	“	40/85	“	“	“

H25c	“	4000	“	1067/2260	“	40/85	“	“	“

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

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

L15	“	1500	“	689/1460	“	40/85	“	“	“

M15	“	2500	“	1067/2260	“	40/85	“	“	“

H15	“	4000	“	1067/2260	“	40/85	“	“	“

Yellow background for Reduced Matrix

The reduced test matrix for fixed compressors included selected 4000 rpm
points.  Each fixed compressor had its speed adjusted to match capacity,
on average, at I35a, L35a, and M35a.

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

900	900	1485

1500	1500	2475

2500	2500	4125

4000	4000	6600

Annualization procedure:

For each individual technology, the annualization weighting factors were
adjusted based on the points that were run for each evaluation.  In
running the final Best-of-Best evaluation with a full test matrix, it
was discovered that the absolute annual savings was over estimated as a
result of doing this evaluation utilizing the reduced test matrix. 
However, the relative ranking of the technologies should still be valid.

Table 5 Ambient weighting profile based on Visteon Climate Model

Outdoor load condition

Temp/ humidity	Weighting for US- climate

<15°C	10.6%

15°C/ 80%	46.2%

25°C/ 50% low	14.2%

25°C/ 50% high	14.2%

25°C/ 80% high	9.5%

35°C/ 40%;low	1.3%

35°C/ 40%;high	3.8%

45°C/ 25%	0.2%

Ignore I60 "a" and "c"	89.4%

Weight elevated idle 33% [only I60b]

	Non elevated 67%

	

Table 6 Compressor Speed profile

Compressor RPM	Percent Time at RPM

	USA

<1000	17%

1500	43%

2500	35%

>4000	5%

System Setup Test Procedures:

Inlet air velocity and temperature distribution:

These tests were performed prior to starting the refrigeration system
performance test to verify uniformity of air velocity and temperature at
the inlet to the heat exchangers.  Each test was performed for both heat
exchangers (condenser and evaporator) and over a relevant range of air
velocities and inlet air temperatures. 

For each of the system tested, the adjustment of air velocity
distribution was not straightforward.  After several iterations,
condenser/gas cooler velocity profiles were achieved within ±8% for
each of the predefined three airflow conditions and the evaporator
velocity profiles were measured within ±6% for each of predefined two
airflow conditions.  The results at selected airflow conditions are
presented below. 

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

Evaporator inlet air temperatures

Evaporator inlet air velocities

Condenser inlet air temperatures

Condenser inlet air velocities

Air side pressure drop:

Each time there was a heat exchanger change, the dry air pressure drop
was measured at room conditions to verify the specification.  The
airside pressure drop of both the evaporator and the condensers are
shown below.  

Evaporator air side pressure drop comparison

Comparison of air side pressure drop of condensers

[Note: Condensers #1/5 have the same air side design as #3]

Refrigerant charge determination:

Charge determination was conducted at three different conditions (I35,
L35 and I25) from the test matrix presented later in this chapter.
Evaluation at three conditions allowed evaluation of optimum charge at
low and high ambient conditions. In some cases only one condition was
run to check charge where it was expected that there would be no impact
on charge.  Optimum charge was picked based on standard industry
practices.

SAE IMAC charge determination test matrix

Test Name	Ambient temp. [oC/oF]	Compr speed [rpm]	Condenser	Evaporator

	Inlet air temp.      [oC/ oF]	Flow rate [l/s/CFM]	Inlet air temp.     
[oC/ oF]	Flow rate [l/s/CFM]	Simulated air selection	RH [%]	Dew point
[oC]

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

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

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

The charging procedure was as follows: after evacuation, the system was
charged with the lowest reasonably expected quantity of refrigerant
500~600g. The system was maintained at constant air temperature,
humidity, and airflow rates through condenser and evaporator, and
compressor speed, as prescribed by the charging condition. When steady
state conditions were achieved data were taken and refrigerant was added
in increments of 50g (or 25g).  An example figure below shows charge
determination test results for one of the systems tested.

Oil circulation rate measurement:

Before the final performance test, oil circulation rate was measured
with two different compressor speeds and conditions.  The oil type, the
initial oil charge, and the target oil circulation values were
recommended by the compressor manufacturer.  Oil circulation was
measured by taking refrigerant/oil samples from the liquid line and
measuring the mass ratio of refrigerant to oil. 

The measurement results for oil circulation rate and system charge
amounts for each system are given in the following table. 



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

900 rpm	4000rpm

	Improved Condensers

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

Condenser #6	1200	-	-	No oil circulation rate were measured for  cond #6
and #7 but with same oil charge 137 cc as the baseline

Condenser #7	1200	-	-	Oil charge 137 cc

Improved Evaporators

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

Evaporator #2	950	-	-	Oil charge 137 cc

Evaporator #3	1050	-	-	Oil charge 137 cc. 

Evaporator #4	1025	-	-	Oil charge 137 cc

Internal Heat Exchangers

Baseline	1100	0.33%	0.87%	comp#2

Commercial 1 HX #1	1150	0.82%	-

	Commercial 2 HX in Parallel #2	1250	1.05% (1500rpm)	-

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

Automotive Tube Type #4	1300	1.84	3.85 (2500rpm)	Oil charge 137 cc,
comp#4

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

Cycling Fixed Compressors:

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

Fixed Piston comp#8	1150	0.93	1.25 (2500rpm)

	Fixed Scroll comp#7	1150	8.2

(1485 rpm)	-

	Insulated Suction Lines

Tested in BoB System with IHX Type #4	1325	-	-	Oil charge 137 cc, comp#4

Improved variable compressors

Baseline comp#1	1350	0.73	1.22

	Compressor #2	1350	0.23	0.87	Integrated oil separator

Compressor #3	1350	0.41	1.42

	Compressor #4	1150	1.46	5.95

	Compressor #6	1150	1.52	2.92

	Condenser Subcooling Pass Designs:

Design #1	1350	0.73	1.22	External receiver

Design #3	1250	‘-	‘-	Integrated receiver

Design #5	1300	‘-	‘-	Integrated receiver

Best of Best System

ARCRPII Baseline	1125	0.66	1.16

	IMAC BoB	1150	1.46	5.95

	Capacity and Energy Test Results:

Improved Condensers:

Comparison of IMAC heat exchangers to industry average:

The OEM team members developed a condenser component specification and
invited condenser manufacturers to submit designs.  The specification
allowed for a thicker core than baseline so long as air side pressure
drop specification was not exceeded.

Condenser designs were submitted by four suppliers.  The OEM members
selected the two highest performance designs including one that is
thicker than the baseline.

SAE IMAC system component table for condenser study

Served as baseline

Component	Feature	Cond#3	Cond#6	Cond#7

Condenser	Core effective face [mm]	W626xH386XD16	W580xH418xD16
W546xH396xD22

	Face area [mm2]   	241,636	242,440	216,216

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

	Aspect ratio [-]	0.62	0.72	0.73

	No. of tubes	39 (micro-channel)	42 (micro-channel)	61 (micro-channel)

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

	Mass [kg]	2.1	2.5	2.5

Evaporator (Evap#3)	Core effective face [mm]	W267xH257xD50

	Face area [mm2]   	68,619

	Core volume [mm3]	3,430,950

	Aspect ratio [-]	0.96

	Type (No. of tubes)	micro-channel

	No. of Passes	 

	Mass [kg]	1.6 (est.)

Compressor (Comp#4)	Type	External controlled variable displacement

	Displacement [cc]	162

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

Receiver/Drier	Mass [kg]	IRD, see condenser

Total Mass	[kg]	10.2	10.6	10.6

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

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

	 

Following are the results of the condenser assessments:

Only the following points were run:

I45, L45, I35a, M35a, I25a, M25a 

Weighting was adjusted to weight only these points.  Weighting of 15 C
was assigned to 25 C.

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

Compr Energy	COP weighted	Compr Energy	COP weighted	Compr Energy	COP
weighted

kJ	 	kJ	 	kJ	 

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

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

Improved Evaporators:

The OEM team members developed an evaporator component specification and
invited evaporator manufacturers to submit designs.  The specification
allowed for a thicker core than baseline so long as air side pressure
drop specification was not exceeded.  

Two suppliers submitted evaporator designs.  Since there were only two
submissions, the OEM members decided to test both including one that is
thicker than the baseline.



SAE IMAC system component table for evaporator study

Served as baseline

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

Condenser (Cond#3)	Core effective face [mm]	W626xH386XD16

	Face area [mm2]   	241,636

	Core volume [mm3]	3,866,176

	Aspect ratio [-]	0.62

	No. of tubes	39 (micro-channel)

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

	Mass [kg]	2.1

Evaporator	Core effective face [mm]	W266xH251xD38	W270xH235xD35
W267xH257xD50

	Face area [mm2]   	66,766	63,450	68,619

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

	Aspect ratio [-]	0.94	0.87	0.96

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

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

	Mass [kg]	1.2	1.1	1.6 (est.)

Compressor (Comp#4)	Type	External controlled variable displacement

	Displacement [cc]	162

	Mass (including pulley) [kg]	6.4

Receiver/Drier	Mass [kg]	IRD, see condenser

Total Mass	[kg]	9.7	9.6	10.1

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

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

	 	 	 	 

Only the following points were run:

I35a, M35a, I25a, M25a

Weighting was adjusted to weight only these points.  Weighting of 15 C
was assigned to 25 C.  Weighting of 45 C was assigned to 35 C.

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

Compr Energy	COP weighted	Compr Energy	COP weighted	Compr Energy	COP
weighted

kJ	 	kJ	 	kJ	 

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

 	 	-7.8%	12.5%	-14.1%	21.2%	COP comparison:

Internal Heat Exchangers w/ superheat optimization:

In ARCRP, traditional co-tube designs were tested with R134a and they
showed little improvement.  The limitation is limited heat exchanger
effectiveness combined with added suction line pressure drop.  Also in
ARCRP, a commercial design was tested with an alternative refrigerant
and it showed more promise.  For I-MAC, the team first tested a single
commercial IHX and two commercial IHXs in parallel for proof of concept.

IHX Study one:

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

Served as baseline

Component	Feature	without IHX 	IHX#1	IHX#2

Condenser (Cond#3)	Core effective face [mm]	W626xH386XD16

	Face area [mm2]   	241,636

	Core volume [mm3]	3,866,176

	Aspect ratio [-]	0.62

	No. of tubes	39 (micro-channel)

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

	Mass [kg]	2.1

Evaporator (Evap#1)	Core effective face [mm]	W266xH251xD38

	Face area [mm2]   	66,766

	Core volume [mm3]	2,537,108

	Aspect ratio [-]	0.94

	No. of tubes	39 (micro-channel)

	No. of Passes	two slabs, passes unknown

	Mass [kg]	1.2

Compressor (Comp#2)	Type	External controlled variable displacement

	Displacement [cc]	160

	Mass (including pulley) [kg]	6.1

Receiver/Drier	Mass [kg]	IRD, see condenser

IHX	Mass [kg]	N/A	1.8	2.8

Total Mass	[kg]	9.4	11.2	12.2

Note:	1. Two IHXs were tested with this system.

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

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

 

Only the following points were run:

L35a, M35a, I25a, M25a, I15, M15

Weighting was adjusted to weight only these points.  Weighting of 45 C
was assigned to 35 C.

Baseline	Baseline	1 IHX	1 IHX	2 IHX	2 IHX

Compr Energy	COP weighted	Compr Energy	COP weighted	Compr Energy	COP
weighted

kJ	 

	kJ	 

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

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

IHX Study two:

Test results for the two commercial IHXs in parallel were good, so the
team decided to test two designs specifically developed for automotive
application.  One is an enhanced co-tube design.  The second is a
compact heat exchanger design.

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

	 	 	Served as baseline

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

Condenser (Cond#7)	Core effective face [mm]	W546xH396xD22

	Face area [mm2]   	216,216

	Core volume [mm3]	4,756,752

	Aspect ratio [-]	0.73

	No. of tubes	61 (micro-channel)

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

	Mass [kg]	2.5

Evaporator (Evap#3)	Core effective face [mm]	W267xH257xD50

	Face area [mm2]   	68,619

	Core volume [mm3]	3,430,950

	Aspect ratio [-]	0.96

	No. of tubes	micro-channel

	No. of Passes	two slabs, passes unknown

	Mass [kg]	1.6 (est.)

Compressor (Comp#4)	Type	External controlled variable displacement

	Displacement [cc]	162

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

Receiver/Drier	Mass [kg]	IRD, see condenser

IHX	Mass [kg]	N/A	2.8	0.4	0.8

Total Mass	[kg]	10.5	13.4	11	11.4

Note:	1. Three  IHXs were tested with this system.

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

	3. IHX#3 was coaxial (0.4 kg) and IHX#4 was core type (0.8 kg).

No real benefit was noted with IHX study two, but this work was done
with a more effective evaporator and the OEM team determined that this
was the major reason that minimal benefits were observed in this study. 
Clearly, the OEM team believes IHX benefits depend on the evaporator
design.

Insulated Suction Line:

As part of the BoB system test, insulating the suction line was
evaluated at select screening points and for this system showed no
significant improvement as illustrated below using the IHX line.

Improved Variable Compressors:

Initially, two compressor suppliers submitted improved designs that were
tested against the baseline compressor.  While an improvement was
measured, the team felt further improvement should be possible.  The two
suppliers were asked to supply second generation design.  Both supplied
newer designs that were also tested.  

Improved Variable Compressor Study one:

The results shown below are for the first two compressors which were
reported out in the December, 2005 Interim report.

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

 	 	Served as baseline	 	 

Component	Feature	Comp#1	Comp#2	Comp#3

Condenser (Cond#5)	Core effective face [mm]	W626xH386XD16

	Face area [mm2]   	241,636

	Core volume [mm3]	3,866,176

	Aspect ratio [-]	0.62

	No. of tubes	39 (micro-channel)

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

	Mass [kg]	2.1

Evaporator (Evap#1)	Core effective face [mm]	W266xH251xD38

	Face area [mm2]   	66,766

	Core volume [mm3]	2,537,108

	Aspect ratio [-]	0.94

	No. of tubes	39 (micro-channel)

	No. of Passes	two slabs, passes unknown

	Mass [kg]	1.2

Compressor	Type	External controlled variable displacement	External
controlled variable displacement	External controlled variable
displacement

	Displacement [cc]	~160	~160	~160

	Mass (including pulley) [kg]	6.4	6.1	6.5

Receiver/Drier	Mass [kg]	IRD, see condenser

Total Mass	[kg]	9.7	9.4	9.8

Note:	1. Three compressors were tested with this system.

	2. EXV as expansion device

	 	 	 	 

Annual results for the 1st round of improved compressors are reported
below with the 2nd test.Improved Variable Compressor Study two:

The results shown below are for the final two compressors.  The
condenser was changed from condenser #5 to condenser #3 for these tests.

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

 	 	Served as baseline

Component	Feature	Comp#1	Comp#4	Comp#6

Condenser (Cond#3)	Core effective face [mm]	W626xH386XD16

	Face area [mm2]   	241,636

	Core volume [mm3]	3,866,176

	Aspect ratio [-]	0.62

	No. of tubes	39 (micro-channel)

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

	Mass [kg]	2.1

Evaporator (Evap#1)	Core effective face [mm]	W266xH251xD38

	Face area [mm2]   	66,766

	Core volume [mm3]	2,537,108

	Aspect ratio [-]	0.94

	No. of tubes	39 (micro-channel)

	No. of Passes	two slabs, passes unknown

	Mass [kg]	1.2

Compressor	Type	External controlled variable displacement	External
controlled variable displacement	External controlled variable
displacement

	Displacement [cc]	~160	~160	~160

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

Receiver/Drier	Mass [kg]	IRD, see condenser

Total Mass	[kg]	9.7	9.8	9.2

Note:	1. Three compressor were tested with this system.

	2. EXV as expansion device

	 

Improved Variable Compressor COP and Energy Annualized Weighting:

COP:

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

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

I45	2.07	2.37	2.31	2.31	2.49	2.59

L45	1.92	2.10	2.06	2.01	2.05	2.31

I35a	2.96	3.04	3.03	2.96	3.16	3.41

L35a	2.60	2.65	2.54	2.57	2.51	2.93

M35a	1.88	1.97	1.91	1.91	1.87	2.15

I25a	3.40	3.47	3.29	3.44	3.60	3.92

L25a	3.07	3.63	3.72	3.24	3.60	3.52

M25a	2.48	3.06	3.40	2.66	3.24	2.98

I15	3.22	3.88	3.72	3.49	4.02	3.90

L15	1.90	2.13	1.86	2.17	3.44	2.67

M15	1.20	1.33	1.41	1.29	2.50	1.38

Weighted:	2.07	2.38	2.38	2.22	2.90	2.51

% Improved compared to Comp#1/Cond#5	14.90%	14.90%	7.16%	39.77%	21.30%

Energy:

Energy Comparison

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

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

I45	1.763	1.539	1.573	1.580	1.465	1.407

L45	2.585	2.367	2.408	2.468	2.420	2.151

I35a	1.531	1.494	1.495	1.534	1.436	1.331

L35a	2.344	2.298	2.393	2.373	2.427	2.076

M35a	3.626	3.461	3.583	3.570	3.650	3.179

I25a	1.190	1.169	1.232	1.177	1.127	1.034

L25a	0.904	0.765	0.746	0.858	0.771	0.789

M25a	1.097	0.888	0.798	1.021	0.838	0.911

I15	0.190	0.158	0.164	0.175	0.152	0.157

L15	0.234	0.209	0.240	0.205	0.129	0.166

M15	0.337	0.305	0.287	0.314	0.162	0.292

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

% change compared to Comp#1/Cond#5	-11.31%	-12.01%	-4.84%	-18.29%
-15.04%

Condenser Sub-cooling Pass Design:

At the beginning of the IMAC project, team two solicited condenser
samples with varying subcooling pass configurations that are illustrated
below.  These condensers were designed to enable use to evaluate the
effects of these configurations on system efficiency.  Five
configurations were submitted, but only the extremes were evaluated.

The total number of tubes for each condenser was 39.  Each heat
exchanger had two passes plus sub-cooler except condenser #1 which does
not have a sub-cooling pass.  Number of tubes in each pass is shown in
the square.  Arrow indicates the flow direction.  Refrigerant circuiting
may not be totally optimized but it is representative of these design
types. 

Condenser 1

 

Condenser 3

 

Condenser 5

 

Three condenser circuiting designs were evaluated.   

SAE IMAC system component table for subcooling study

Served as baseline

Component	Feature	Cond#1	Cond#3	Cond#5

Condenser	Core effective face [mm]	W626xH386XD16	W626xH386XD16
W626xH386XD16

	Face area [mm2]   	241,636	241,636	241,636

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

	Aspect ratio [-]	0.62	0.62	0.62

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

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

	Mass [kg]	1.8	2.1	2.1

Evaporator (Evap#1)	Core effective face [mm]	W266xH251xD38

	Face area [mm2]   	66,766

	Core volume [mm3]	2,537,108

	Aspect ratio [-]	0.94

	No. of tubes	39 (micro-channel)

	No. of Passes	two slabs, passes unknown

	Mass [kg]	1.2

Compressor (Comp#1)	Type	External controlled variable displacement

	Displacement [cc]	160

	Mass (including pulley) [kg]	6.4

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

Total Mass	[kg]	9.8	9.7	9.7

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

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

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

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

SAE IMAC Condenser Circuiting

	Number of tubes	Sub-cooling to total surface [%]

	First pass	Second pass	Sub-cooling pass

	Condenser #1	26	13	0	0

Condenser #3	22	11	6	18

Condenser #5	18	9	12	44

COP vs subcooling

Capacity vs Subcooling

COP vs Charge

Sub-cooling Control Strategy:

To evaluate the effect of varying sub-cooling with the same condenser, a
by-pass configuration was devised to allow liquid to separate in the
receiver and by-pass the sub-cooling loop.  Additional valves were added
to attempt to take the gas/liquid from the top of the receiver and route
this to the inlet of the sub-cooling pass to take better advantage of
the heat exchange capability of this part of the condenser at low loads.
 Below is an illustration of this plumbing.  The results were less than
satisfactory and will be discussed later in the report.  It is believed
that the liquid separation may not have been adequate to provide the
expected advantage.

The evaluation of controlled sub-cooling system did not provide a
significant improvement in COP.  However, after review of the data in
more detail, the team feels that the receiver may not have separated the
liquid well during the test and improved hardware may produce more
promising results.  

COP and Qe results for subcooling control

The results above indicate that there are minimal benefits from the
setup and operating strategy used.  Team 2 feels that there still may be
merit but a better method of separating the gas from the liquid in the
receiver will be necessary to better evaluate this technology.   

Since there was not significant improvement noted at these two screening
points an annual number was not calculated.

Superheat Control Strategy:

An Electronic Expansion Valve was used.  The expansion valve controlled
superheat at equal capacity and measured the effect on system COP.  This
study was done first on the baseline 38 mm. thick evaporator and then
later on the optimized 50 mm thick evaporator.

Superheat control study one:

SAE IMAC system component table for superheat study

Component	Feature	 

Condenser (Cond#5)	Core effective face [mm]	W626xH386XD16

	Face area [mm2]   	241,636

	Core volume [mm3]	3,866,176

	Aspect ratio [-]	0.62

	No. of tubes	39 (micro-channel)

	No. of Passes	2 (26-13)

	Mass [kg]	2.1

Evaporator (Evap#1)	Core effective face [mm]	W266xH251xD38

	Face area [mm2]   	66,766

	Core volume [mm3]	2,537,108

	Aspect ratio [-]	0.94

	No. of tubes	39 (micro-channel)

	No. of Passes	two slabs, passes unknown

	Mass [kg]	1.2

Compressor (Comp#2)	Type	External controlled variable displacement

	Displacement [cc]	160

	Mass (including pulley) [kg]	6.1

Receiver/Drier	Mass [kg]	IRD, see condenser

Total Mass	[kg]	9.4

Note:	 

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

	 

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

This result is largely due to the poor temperature distribution of the
evaporator being used and the result may be different with other
evaporators.

Baseline-7K	Baseline-7K	4K SH	4K SH	1K SH	1K SH

Compr Energy	COP weighted	Compr Energy	COP weighted	Compr Energy	COP
weighted

kJ	 

	kJ	 

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

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

Superheat control study two:

Best of the Best System:

The Baseline for comparison is the R134a system from ARCRPII that was
run to normal evaporator air off temperatures.

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

B-o-B Improved Condenser

B-o-B Improved Evaporator

B-o-B Improved External Control Variable Compressor 

Superheat Control Algorithm w/ EXV

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

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

Component	Feature	B-o-B - without IHX 

Condenser (Cond#7)	Core effective face [mm]	W546xH396xD22

	Face area [mm2]   	216,216

	Core volume [mm3]	4,756,752

	Aspect ratio [-]	0.73

	No. of tubes	61 (micro-channel)

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

	Mass [kg]	2.5

Evaporator (Evap#3)	Core effective face [mm]	W267xH257xD50

	Face area [mm2]   	68,619

	Core volume [mm3]	3,430,950

	Aspect ratio [-]	0.96

	No. of tubes	micro-channel

	No. of Passes	two slabs, passes unknown

	Mass [kg]	1.6 (est.)

Compressor (Comp#4)	Type	External controlled variable displacement

	Displacement [cc]	~160

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

Receiver/Drier	Mass [kg]	IRD, see condenser

IHX	Mass [kg]	N/A

Total Mass	[kg]	9.4

The numbers have been annualized a number of ways.  First, as compared
to the ARCRP Phase II R134a system and then as compared to the original
baseline system that was assembled for the IMAC project.  For the
comparison below, the data was analyzed with equal evaporator air outlet
temperatures and the capacity was averaged for the energy calculation to
assure comparable results.  Then the tests were rerun with the
evaporator out air temperature set to 10°C and the comparison in
results is shown below.

ARCRP Base	IMAC B-o-B

Teao= ARCRP	IMAC B-o-B

Teao= 10C

Compr Energy	COP weighted	Compr Energy	COP weighted	Compr Energy	COP
weighted

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

-12.3%	15.0%	-36.4%	12.6%

The IMAC baseline did not run the complete test matrix, so this
analysis is done with a reduced test matrix. [I45, L45, I35a, L35a,
M35a, I25a, L25a, M25a, I15, L15, M15]  You can see that the IMAC
baseline was less efficient than the original ARCRP reference.  But the
improvement in COP is significant as compared to the previous
comparison.  The improvement in energy consumption is slightly smaller,
but all of these tests were run with the evaporator air outlet
temperature equal to 10°C.

IMAC Base Compr energy	IMAC B-o-B Compr energy	IMAC Base COP	IMAC B-o-B
COP

663,495	481,791	2.20	3.39

	27.4%

54.3%

Fixed vs Variable Compressors:

The OEM members selected two production fixed displacement compressors
for test.  One is a so-called High Efficiency Piston design and the
other is a Scroll design.  The manufacturers were asked to supply sample
compressors.  Both suppliers provided compressors and assistance
specifying oil type and circulation rates.  The baseline compressor for
this test is the BoB variable.  The OEM team believed that a cycling
system would behave differently with an EXV compared to a TXV.  It was
decided to conduct this testing with a TXV.  The TXV was specified by
the evaporator supplier.  The baseline variable compressor was run first
with the TXV.

SAE IMAC system component table for fixed compressor study

 	 	Served as baseline

Component	Feature	comp#4 	comp#7	comp#8

Condenser (Cond#7)	Core effective face [mm]	W546xH396xD22

	Face area [mm2]   	216,216

	Core volume [mm3]	4,756,752

	Aspect ratio [-]	0.73

	No. of tubes	61 (micro-channel)

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

	Mass [kg]	2.5

Evaporator (Evap#3)	Core effective face [mm]	W267xH257xD50

	Face area [mm2]   	68,619

	Core volume [mm3]	3,430,950

	Aspect ratio [-]	0.96

	No. of tubes	micro-channel

	No. of Passes	two slabs, passes unknown

	Mass [kg]	1.6 (est.)

Compressor	Type	External controlled variable displacement	Fixed
displacement scroll type	Fixed displacement piston type

	Displacement [cc]	~160	~90	~160

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

Receiver/Drier	Mass [kg]	IRD, see condenser

Total Mass	[kg]	10.6	8.6	9.7

Note:	1. Three compressors were tested with this system.

	2. TXV as expansion device

	 



Annualization:

IMAC EXV	IMAC TXV	Fixed Compr #7	Fixed Compr #8	IMAC EXV	IMAC TXV	Fixed
Compr #7	Fixed Compr #8

Compr Energy	Compr Energy	Compr Energy	Compr Energy	COP weighted	COP
weighted	COP weighted	COP weighted

 	kJ	 	 	 	 	 	 

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

0%	base	20%	27%	-3%	base	-11%	-19%

III. Vehicle Testing

A Cadillac STS was used as a vehicle test bed to demonstrate that an
improved efficiency refrigerant system would provide equal cooling
capacity.  The plan also included development of control strategy
algorithms.  

Demonstration of equal capacity:

The main objective of the test vehicle was to demonstrate equal AC
performance.  This was proved during the Phoenix Alternative Refrigerant
System Symposium (shown below) during the ride/drive evaluation (varying
city traffic / Idle driving pattern). 



The vehicle development plan is shown below which progressed with 3
generations of algorithm improvement, with a final verification of the
3rd/ final generation.

Vehicle Build and Test Plan

 

Generation One:

The vehicle was built with an optimized external controlled variable
compressor, an electronic expansion valve, D-Space controller, MATLAB
and SIMULINK software which used a new algorithm for control of the
compressor and expansion valve to optimized COP control.  All other
components were original equipment parts from the 2005 Cadillac STS. 

The first generation of software was produced by EDS from GM based
software.  The algorithm was designed to operate in two phases.  The
first was a pull down region where the cabin temperature is greater than
the desired temperature.  In this region the compressor displacement was
controlled to the maximum displacement, the expansion valve was
controlled to maintain optimal superheat for maximum capacity, and the
air recirculation door was set to 0% outside air.  The second phase,
(comfort region) took over when the desired cabin temperature has been
reached, and the capacity could be reduced to minimize energy usage,
while maintaining occupant comfort.  In the comfort region, the
compressor displacement is reduced as much as possible to raise the
evaporator outlet pressure keeping the evaporator temperature as high as
possible consistent with vehicle occupant comfort.  The recirculation
door is opened periodically to allow for sufficient outside air for
occupant comfort, and the EXV is controlled to maintain an optimal
superheat for maximum COP.

The vehicle was fully instrumented and outfitted with several components
to achieve the desired control and measure the effects.  A D-Space
MicroAutobox is used along with a driver-box to read the parameters
valid for control, run the control algorithm, and output the desired
control signals to the compressor and EXV controller.  The following
diagram shows the vehicle hardware setup.

The software model was implemented in MatLab Simulink using the dSpace
RTI library and executes on the MicroAutobox controller.  The software
model is divided into four layers, the Hardware Input Output (HWIO)
Layer which reads the data inputs and provides the controller outputs,
the Software Integration (SWIL) Layer which handles general services,
the Application Layer which composes the main body of the application,
and the Operating System (OS) Layer which provides controller
management.

), suction density (ρs) and volumetric efficiency (ηv) as shown in the
equation below.  The duty cycle was estimated based on the control
signal to the compressor and the volumetric efficiency was estimated
based on discharge pressure.  Using the mass flow rate from this method
predicted unreasonably high capacities originally.

It was also determined that the original expansion control valve driver
was too slow and that the hardware/sensors needed to be calibrated. The
data acquisition system (DAS) was also overheating which need to be
resolved.   

Creative Thermal Solutions (CTS) in Urbana, IL was chosen to debug the
vehicle and algorithm to provide additional functionality to the A/C
system control and improve the algorithm.     

CTS developed a new electronic expansion valve driver control board and
software, to be able to speed up the superheat adjustment, they resolved
the DAS overheating, calibrated the pressure transducers, made
corrections to the COP and normalized load calculation, and re-wired the
Micro Auto box inputs.   

To evaluate these changes the vehicle was taken on a test drive, 
however after some corrections to the implementation of this mass flow
rate (M r) equation the prediction preformed better as is shown in the
capacity plot as ‘EDS method corrected’.

  .

The corrected calculation (found units errors and multiplying by
specific volume instead of dividing by it) gave reasonable capacity
values at the full displacement I25b point but performed poorly at
reduced displacement points because there was no accounting for the loss
in volumetric efficiency at reduced displacements and because the duty
cycle did not directly correlate to the displacement.  The torque was
also estimated using a normalized load calculation which resulted in
high predictions of compressor work.

  

 It was determined during the test drive from Urbana, IL to Laredo, TX.,
that the algorithm still needed further improvement to calculate COP and
capacity.   The link between efficiency and displacement was difficult
to separate and estimate with reasonable accuracy. 

The following plot shows the effect of changes on compressor speed and
evaporator air exit temperature (EAT) on COP.  While the algorithm
predicted COP values in a reasonable range for part of the test, clearly
the prediction was not reasonable over the entire test.   

Generation two:

Following the Texas ride, the vehicle was returned to Detroit to switch
compressors to the compressor that performed best at the University of
Illinois ACRC labs.  At this time the method (2nd Algorithm version) for
estimating mass flow rate and compressor power was changed to use linear
regression.  The regression was based off of the air flow rate on the
condenser (AFRc), the air temperature entering the condenser (Tcai), the
air temperature exiting the evaporator measured at the nozzle (Ten), the
compressor speed (Vc), the discharge pressure (Pcpro), the compressor
displacement control current (CompCurr) and the air flow rate over the
evaporator (AFRe).  

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

A regression for compressor power was done in the same fashion
calculating a different term (Kp) using the equation below.

This model was found to match a limited number of data points that it
was based on reasonably well as can be seen in the following plots
showing bench data vs. regression predictions for refrigerant mass flow
rate and compressor power.  However, when the regressions were used in
the vehicle at conditions differing from the bench conditions the mass
flow rate and compressor work predicted values 2 to 3 times higher than
expected. 

The vehicle was taken to Phoenix in the summer, 2006 and high
temperature testing was performed.  Interestingly, the effect of
superheat on capacity and COP was found to be different than had been
previously observed on the bench.  During idle, the optimal superheat
for maximum COP and capacity was observed to be 10°C to 15°C whereas
bench data has indicated that superheats near 0°C are optimal.  In
order to further understand the effect of superheat on COP and capacity,
vehicle testing was performed in the climatic chamber facility to
eliminate dynamics that may have affected these results.

Generation three:

After the Phoenix Alternative Refrigerant System Symposium, the
Algorithm was then modified (3rd Version), to incorporate a physically
based approach.  In this approach the original mass flow rate equation
which calculated mass flow rate by multiplying compressor displacement,
speed, suction density and volumetric efficiency was used.  The
compressor displacement and volumetric efficiency were combined into one
term (Kv) and the regression was performed basing this term on a
combination of pressure ratio (Pd/Ps), compressor pressure difference
*(Pd-Ps), compressor speed (Vc), and compressor displacement control
signal current (CompCurr).  The regression equation (physically based
regression) shown below was used with the full ACRC test matrix data to
calculate the constants C1 through C7.

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

A regression for compressor power was done in the same fashion
calculating a different term (Kp) using the equation below.  

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

The results from this regression fit the ACRC test bench data very well
as can be seen in the following plots showing the regression vs. bench
data for COP, capacity, and compressor power.  It should be noted that
the regression predicts high values of COP.

The vehicle was tested in the climatic chamber facility at CTS to match
4 conditions from the bench test matrix (conditions selected I35b, L35b,
I25b, and L25b).  The results matched the bench data very well showing
good performance from the regression.  The following plots show the
vehicle results compared to the bench data for COP, capacity and
compressor power.  It should be noted that for L35b the compressor
displacement was higher than on the bench resulting in a higher capacity
and compressor power.

After returning from Phoenix it was decided to study the effect of
superheat, and its relationship to COP, and Capacity in the vehicle to
determine why the 2nd Algorithm version gave conflicting results from
Bench testing.  The study also examined the effects of re-circulated air
control and evaporator temperature control on compressor work.  The test
points in the vehicle were at IMAC conditions I25b, L25b, and M25b as
well as a condition matching the Phoenix point (45C, 1000 engine rpm). 
At each condition, the vehicle was tested at superheats of 2°C, 5°C,
10°C, 15°C, and 20°C while all other parameters were maintained
constant.  There was difficulty maintaining the superheat constant at
the low displacement conditions, particularly at L25b and M25b.  System
instability was observed to be the most severe at superheats of 10°C
and 15°C.  The following plot shows the superheat instability at
various superheat set points which could be eliminated given more time
to calibrate the expansion valve Proportional Integral Derivative (PID)
constants.   

 

Shown below is an example of the Proportional and Derivative constant
tuning to give better control of superheat.

A refrigerant mass flow meter was installed prior to running the
superheat study so the capacity could be calculated using this flow rate
multiplied by the refrigerant enthalpy change across the evaporator.  

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

The following plot shows that for all of the operating conditions tested
the capacity drops as superheat increases and that a superheat of 2°C
provide an optimum capacity.  During all the superheat variation tests
for each condition, the compressor control signal was set to a constant
value and only the expansion valve setting was varied.  

Effect of evaporator Superheat on COP at constant capacity:

COP now behaves similar to the test bench.  The I25b and Phoenix
operating conditions were at full displacement and showed that COP was
highest at the lowest superheats.  At the L25b and M25b conditions the
compressor displacement was adjusted to maintain a constant capacity
(2.4kW).  

Effect of varied evaporator air set point:

An optimal superheat of 2°C was chosen for the four operating
conditions, and EAT was varied by reducing compressor displacement to
simulate capacity reduction potential once the cabin of the vehicle was
cooled.  The EAT was reduced from approximately 3°C to 14°C for the
conditions with the starting EAT depending on the lowest that could be
achieved at a given condition.  For the Phoenix condition the lowest EAT
attainable was higher than 14°C so the condition was left out of the
EAT study.  The following plots show the capacity, compressor power
reduction, and COP increases attainable by allowing the EAT to increase.
 The COP increases would be less then the estimated COP’s show here
due to the fact that COP is predicted somewhat high at low displacements
as shown previously.  Significant compressor power savings can be
achieved by reducing displacement after the cabin of the vehicle is
cooled and Evaporator set-point logic is invoked.

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

A study to determine the power savings attainable by using higher
amounts of recirculation air was also performed.  For each of the
conditions a range of recirculation air percentage from 0% to 100% was
tested.  The following plots show significant compressor power and
capacity savings can be attained by increasing the air recirculation
amounts.  Furthermore at the I25b and Phoenix conditions the compressor
was operating at full displacement thus much lower evaporation
temperatures were attainable at full recirculation.

Verification of capacity regression [Measured vs Algorithm]:

Adding the refrigerant flow meter allowed for an accurate measurement of
system capacity and also allowed for a quality check of the physical
based regression algorithm.  The following plot shows a comparison of
the capacity calculated using the mass flow meter and enthalpy change
across the evaporator, and the capacity calculated using the final
physically based regression.   All of the data taken in the superheat,
Evaporator air Temperature [EAT], and recirculation air studies is shown
in the plot.  Compared to the original capacity estimation methods where
the capacity was predicted 2 to 3 times higher than actual this
regression performs very well.  

Summary of the algorithm comparisons:

The plots below show results from the 3 algorithm iterations calculated
with vehicle data for the estimation of COP, capacity, and compressor
power, as well as bench test results at the same conditions, and one
additional method for COP calculation (Enthalpy based), and another
additional method for capacity calculation (Air Side Capacity). This
should help understand the progression / refinement of the algorithm.  
The first iteration was a physically based approach implemented by EDS,
the second was a linear regression approach implemented, and the third
was a physically based regression approach.  

Enthalpy Based COP Calculation

An additional method for COP calculation was implemented using only the
refrigerant enthalpies.  The COP was calculated by dividing the change
in refrigerant enthalpy across the evaporator by the change in
refrigerant enthalpy across the compressor as shown in the equation
below.  A constant, E1, was calculated using the bench data to account
for compressor efficiency. 



Airside Capacity Calculation

A dry air side capacity measurement was also evaluated.  As expected,
this method outputs low values at the four conditions shown in the
capacity plot below due to the lack of the latent load component.  To
calculate the capacities the air flow rate (AFRc) was multiplied by the
specific heat and the change in air temperature across the evaporator as
shown in the equation below.



Conclusion:

The final physically based regression showed a very good agreement with
the bench as did the enthalpy based COP calculation.  Promising results
have been observed quantifying the potential savings in energy usage by
using the control algorithm.  By reducing the compressor displacement
and increasing the EAT after the cabin has been cooled, significant
energy savings can be achieved.  Additional savings can be achieved by
increasing the air recirculation and reducing the superheat.  The
control algorithm has been set to control all of these parameters to
maximize energy savings.

IMAC Team 2 - Improved Efficiency	Final Report		April, 2007

 PAGE   

	Page   PAGE  1  of   NUMPAGES  99 

air mass flow rate

h

ai

=h (T, RH)

h

ao

=h (T, RH)

refrigerant mass flow rate

h

ri

=h (T, P, x)

h

ro

=h (T, P)

±

steam

glycol

air mass flow rate

h

ai

=h (T, RH)

h

ao

=h (T, RH)

refrigerant mass flow rate

h

ri

=h (T, P, x)

h

ro

=h (T, P)

±

steam

glycol

condensate

26

11

22

18

9

2 of 3 evaluators never reach a 

5-comfort (best)

2 of 3 evaluators reach a 5-comfort (best), better than production 

Task

Responsibility

Completion

Comments

Obtain Vehicle

Bill Hill

Complete - see 

Note 1

GM will provide 2 Cadillac STS Vehicles 

(Efficiency and Heatload)

Identify Test Procedures:

Team 2 OEMs

Select Refg Control Parameters

Team 2

Complete

Based on stand test at University of Illinois

Data Acquisition

Bill Hill

Complete

GM will provide 2 systems

Source Refg Sys Controls Alg Dev

Bill Hill/Supplier

Complete

EDS

Parts to Nissan

Suppliers/University of Illinois

Complete

Maintain supplier confidentiality

Install EXV + Data Acquisition

GM

Complete

Install EXV, datalogger, dSpace Controller

Prep & Instrument Vehicles

Dave Barwin

Complete

Nissan to install parts and instrument

Vehicle Algorithm dev

Team 2

Complete

In Detroit

Ship Vehicles to Golden, CO

Bill Hill

Complete

Vehicle Drive Eval - Golden, CO

Bill Hill and Dave Barwin

Complete

Vehicle at NREL - Golden, CO

Bill Hill

Complete

Soak & cool down test for CAE Model Corr

Vehicle Algorithm dev

Team 2

Complete

In Detroit

Wind Tunnel Test

Team 2

Sept 20-21, 2005

Scrubbed due to instrumentation problems

Develop Future Vehicle Test Plans

Team 2

Complete

Dedicated development engineer needed to 

continue the vehicle test plan

Algorithm contract with CTS

CTS / Dave Barwin

Complete

Outsourced development to CTS

Test Drive

OEMs / Dave Barwin

Complete

Install Supplier B compressor

Supplier B

Complete

Additional Algorithm development

CTS / Dave Barwin

Complete

Ship Vehicles to Phoenix

Complete

Prep Drive in Phoenix

OEMs / Dave Barwin

Complete

Sponsor Ride

Complete

Phoenix Symposium

OEMs / Dave Barwin

Complete

Ship Vehicles to NREL

Complete

Solar soak comparison to Team 3

Ship Vehicle to Detroit

Complete

Phase 3 verify Super heat vs Cop

CTS / Dave Barwin

Complete

Completed

Team 2 Efficiency Vehicle Test Plan-Plan C

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ᄀObjective:  Demonstrate that an improved efficiency A/C system meets
normal cooldown performance.

 

 

COP values for low compressor displacements as can be seen in the
conditions M25a, L25b, M25b, L25c, and M25c.