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

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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.


Name:           BRANCH MICHAEL
Title:
Organization:   BRANCH HIGHWAYS INC
Date:           20030310
Address1:
Address2:
City:           ROANOKE
State:          VA
Zip Code:       24022
Docket Number:  S030
Exhibit Number: 7-10
- - - - - - - - - - - - - - - - - - - - - - - - - - - - - - -

Safety Standards for Cranes and Derricks

The Occupational Safety and Health Administration is planning to establish a Crane and Derrick Negotiated Rulemaking Advisory Committee (C-DAC) under the Negotiated Rulemaking Act (NRA) and the Federal Advisory Committee Act (FACA). The Committee will include representatives of parties who would be significantly affected by the final rule.

On July 16, 2002, OSHA published a Federal Register notice of intent to establish a negotiated rulemaking committee (67 FR 46612) and requested nominations for membership. In the current Federal Register notice OSHA is asking for comments on the proposed list of fifty-five nominations.  The list can be found in the complete notice.

Please note that you may not attach materials such as studies or journal articles to your electronic comments. If you wish to include such materials, you must submit three copies to the OSHA Docket Office at the address listed above. When submitting such materials to the OSHA Docket Office, clearly identify your electronic comments by name, date, subject, and Docket Number, so that we can attach the materials to your electronic comments.

Online comments must be received by March 31, 2003 COB EST.  If you wish to commit online on the proposed list of Committee members, please enter your submission in the box below.  Thank you for your interest in worker safety and health.

for Stephen Wiltshire: I would like to express support for his selection to the rulemaking committee.


[Federal Register Volume 80, Number 16 (Monday, January 26, 2015)]
[Proposed Rules]
[Page 3936]
[FR Doc No: 2015-01039]

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ENVIRONMENTAL PROTECTION AGENCY

40 CFR Part 271

[EPA-R04-RCRA-2014-0710; FRL-9921-91-Region 4]

Georgia: Final Authorization of State Hazardous Waste Management
Program Revisions

AGENCY: Environmental Protection Agency (EPA).

ACTION: Proposed rule.

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SUMMARY: Georgia has applied to the Environmental Protection Agency
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DATES: Comments must be received in writing by February 25, 2015.

ADDRESSES: Submit your comments, identified by Docket ID No. EPA-R04-
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     Hand Delivery or Courier: Deliver your comments to
Gwendolyn Gleaton, Permits and State Programs Section, RCRA Programs
and Materials Management Branch, RCRA Division, U.S. Environmental
Protection Agency, Atlanta Federal Center, 61 Forsyth Street SW.,
Atlanta, Georgia 30303-8960. Such deliveries are only accepted during
the Regional Office's normal hours of operation, and special
arrangements should be made for deliveries of boxed information.
    Please see the immediate final rule in the ``Rules and
Regulations'' section of today's Federal Register for detailed
instructions on how to submit comments.

FOR FURTHER INFORMATION CONTACT: Gwendolyn Gleaton, Permits and State
Programs Section, RCRA Programs and Materials Management Branch, RCRA
Division, U.S. Environmental Protection Agency, Atlanta Federal Center,
61 Forsyth Street SW., Atlanta, Georgia 30303-8960; telephone number:
(404) 562-8500; fax number:<PHONE_NUMBER>; email address:
<EMAIL_ADDRESS>
SUPPLEMENTARY INFORMATION: Along with this proposed rule, EPA is
publishing an immediate final rule in the ``Rules and Regulations''
section of today's Federal Register pursuant to which EPA is
authorizing these changes. EPA did not issue a proposed rule before
today because EPA believes this action is not controversial and does
not expect comments that oppose it. EPA has explained the reasons for
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period, the direct final rule in today's Federal Register will become
effective on the date it establishes, and EPA will not take further
action on this proposal. If EPA receives comments that oppose this
action, EPA will withdraw the direct final rule and it will not take
effect. EPA will then respond to public comments in a later final rule
based on this proposed rule. You may not have another opportunity to
comment on these State program changes. If you want to comment on this
action, you must do so at this time. For additional information, please
see the immediate final rule published in the ``Rules and Regulations''
section of today's Federal Register.

    Dated: January 13, 2015.
Heather McTeer Toney,
Regional Administrator Region 4.
[FR Doc. 2015-01039 Filed 1-23-15; 8:45 am]
BILLING CODE 6560-50-P


[Federal Register Volume 78, Number 61 (Friday, March 29, 2013)]
[Notices]
[Pages 19364-19366]
[FR Doc No: 2013-07267]

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DEPARTMENT OF TRANSPORTATION

National Highway Traffic Safety Administration

[Docket No. NHTSA-2013-0016; Notice 1]

Notice of Receipt of Petition for Decision That Nonconforming
1992-1994 BMW 3-Series Passenger Cars are Eligible for Importation

AGENCY: National Highway Traffic Safety Administration, DOT.

ACTION: Notice of receipt of petition.

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SUMMARY: This document announces receipt by the National Highway
Traffic Safety Administration (NHTSA) of a petition for a decision that
nonconforming 1992-1994 BMW 3-Series passenger cars that were not
originally manufactured to comply with all applicable Federal Motor
Vehicle Safety Standards (FMVSS), are eligible for importation into the
United States because they are substantially similar to vehicles that
were originally manufactured for sale in the United States and that
were certified by their manufacturer as complying with the safety
standards (the U.S.-certified version of the same 1992-1994 BMW 3-
Series passenger cars) and they are capable of being readily altered to
conform to the standards.

DATES: The closing date for comments on the petition is April 29, 2013.

ADDRESSES: Comments should refer to the docket and notice numbers above
and be submitted by any of the following methods:
     Federal eRulemaking Portal: Go to http://www.regulations.gov. Follow the online instructions for submitting
comments.
     Mail: Docket Management Facility: U.S. Department of
Transportation, 1200 New Jersey Avenue SE., West Building Ground Floor,
Room W12-140, Washington, DC 20590-0001
     Hand Delivery or Courier: West Building Ground Floor, Room
W12-140, 1200 New Jersey Avenue SE., between 9 a.m. and 5 p.m. ET,
Monday through Friday, except Federal holidays.
     Fax:<PHONE_NUMBER>
    Instructions: Comments must be written in the English language, and
be no greater than 15 pages in length, although there is no limit to
the length of necessary attachments to the comments. If comments are
submitted in hard copy form, please ensure that two copies are
provided. If you wish to receive confirmation that your comments were
received, please enclose a stamped, self-addressed postcard with the
comments. Note that all comments

received will be posted without change to http://www.regulations.gov,
including any personal information provided. Please see the Privacy Act
heading below.
    Privacy Act: Anyone is able to search the electronic form of all
comments received into any of our dockets by the name of the individual
submitting the comment (or signing the comment, if submitted on behalf
of an association, business, labor union, etc.). You may review DOT's
complete Privacy Act Statement in the Federal Register published on
April 11, 2000 (65 FR 19477-78).
    How to Read Comments submitted to the Docket: You may read the
comments received by Docket Management at the address and times given
above. You may also view the documents from the Internet at http://www.regulations.gov. Follow the online instructions for accessing the
dockets. The docket ID number and title of this notice are shown at the
heading of this document notice. Please note that even after the
comment closing date, we will continue to file relevant information in
the Docket as it becomes available. Further, some people may submit
late comments. Accordingly, we recommend that you periodically search
the Docket for new material.

FOR FURTHER INFORMATION CONTACT: Coleman Sachs, Office of Vehicle
Safety Compliance, NHTSA<PHONE_NUMBER>).

SUPPLEMENTARY INFORMATION:

Background

    Under 49 U.S.C. 30141(a)(1)(A), a motor vehicle that was not
originally manufactured to conform to all applicable FMVSS shall be
refused admission into the United States unless NHTSA has decided that
the motor vehicle is substantially similar to a motor vehicle
originally manufactured for importation into and sale in the United
States, certified under 49 U.S.C. 30115, and of the same model year as
the model of the motor vehicle to be compared, and is capable of being
readily altered to conform to all applicable FMVSS.
    Petitions for eligibility decisions may be submitted by either
manufacturers or importers who have registered with NHTSA pursuant to
49 CFR Part 592. As specified in 49 CFR 593.7, NHTSA publishes notice
in the Federal Register of each petition that it receives, and affords
interested persons an opportunity to comment on the petition. At the
close of the comment period, NHTSA decides, on the basis of the
petition and any comments that it has received, whether the vehicle is
eligible for importation. The agency then publishes this decision in
the Federal Register.
    J.K. Technologies, LLC. of Baltimore, Maryland (Registered Importer
90-006) has petitioned NHTSA to decide whether nonconforming 1992-1994
BMW 3-Series passenger cars are eligible for importation into the
United States. The vehicles which J.K. Technologies believes are
substantially similar are 1992-1994 BMW 3-Series passenger cars that
were manufactured for sale in the United States and certified by their
manufacturer as conforming to all applicable FMVSS.
    The petitioner claims that it compared non-U.S. certified 1992-1994
BMW 3-Series passenger cars to their U.S.-certified counterparts, and
found the vehicles to be substantially similar with respect to
compliance with most FMVSS.
    J.K. Technologies submitted information with its petition intended
to demonstrate that non-U.S. certified 1992-1994 BMW 3-Series passenger
cars, as originally manufactured, conform to many FMVSS in the same
manner as their U.S. certified counterparts, or are capable of being
readily altered to conform to those standards. Specifically, the
petitioner claims that non-U.S. certified 1992-1994 BMW 3-Series
passenger cars are identical to their U.S. certified counterparts with
respect to compliance with Standard Nos. 102 Transmission Shift Lever
Sequence, Starter Interlock, and Transmission Braking Effect, 103
Windshield Defrosting and Defogging Systems, 104 Windshield Wiping and
Washing Systems, 105 Hydraulic and Electric Brake Systems, 106 Brake
Hoses, 109 New Pneumatic Tires, 113 Hood Latch System, 116 Motor
Vehicle Brake Fluids, 118 Power-Operated Window, Partition, and Roof
Panel Systems, 124 Accelerator Control Systems, 201 Occupant Protection
in Interior Impact, 202 Head Restraints, 204 Steering Control Rearward
Displacement, 205 Glazing Materials, 206 Door Locks and Door Retention
Components, 207 Seating Systems, 209 Seat Belt Assemblies, 210 Seat
Belt Assembly Anchorages, 211 Wheel Nuts, Wheels Disks, and Hub Caps,
212 Windshield Mounting, 214 Side Impact Protection, 216 Roof Crush
Resistance, 219 Windshield Zone Intrusion, and 302 Flammability of
Interior Materials.
    The petitioner also contends that the vehicles are capable of being
readily altered to meet the following standards, in the manner
indicated:
    Standard No. 101 Controls and Displays: replacement of the
instrument cluster with the U.S.-model component and reprogramming the
vehicle computer.
    Standard No. 108 Lamps, Reflective Devices and Associated
Equipment: replacement of the headlamps, side marker lamps, and tail
lamps with U.S.-model components.
    Standard No. 110 Tire Selection and Rims for Motor Vehicles with a
GVWR of 4,536 kilograms (10,000 pounds) or Less: installation of a tire
information placard.
    Standard No. 111 Rearview Mirrors: replacement of the passenger
side rearview mirror with a U.S.-model component or inscription of the
required warning statement on the face of that mirror.
    Standard No. 114 Theft Protection and Rollaway Prevention:
reprogramming the vehicle computer to activate the key warning and belt
warning systems.
    Standard No. 118 Power-operated Window, Partition, And Roof Panel
Systems: reprogramming the vehicle computer to conform to the standard.
    Standard No. 208 Occupant Crash Protection: reprogramming the seat
belt warning lamp to activate in the proper manner. The petitioner
states that the automatic restraint system in the non-U.S. certified
vehicle complies with the standard and is identical to that found on
its U.S.-certified counterpart, but that the lap and shoulder belts at
the front and rear outboard seating positions must be replaced to
conform to the standard.
    The petitioner states that the bumper shocks must be replaced with
U.S.-model components in order to comply with the Bumper Standard at 49
CFR Part 581.
    The petitioner additionally states that a vehicle identification
plate must be affixed to the vehicles near the left windshield post and
a certification label must be added in the left front door post area to
meet the requirements of 49 CFR Part 565.
    All comments received before the close of business on the closing
date indicated above will be considered, and will be available for
examination in the docket at the above addresses both before and after
that date. To the extent possible, comments filed after the closing
date will also be considered. Notice of final action on the petition
will be published in the Federal Register pursuant to the authority
indicated below.

    Authority:  49 U.S.C. 30141(a)(1)(A), (a)(1)(B), and (b)(1); 49
CFR 593.7; delegation of authority at 49 CFR 1.95 and 501.8.

    Issued on: March 22, 2013.
Claude H. Harris,
Director, Office of Vehicle Safety Compliance.
[FR Doc. 2013-07267 Filed 3-28-13; 8:45 am]
BILLING CODE 4910-59-P