The GreenChill Advanced Refrigeration Partnership
GREENCHILL
     September 2, 2008
 EPA-430-R-08-010
     Theoretical Analysis of Alternative
     Supermarket Refrigeration Technologies
     U.S. Environmental Protection Agency
     Stratospheric Protection Division (6205J)
     1200 Pennsylvania Avenue, NW
     Washington, DC 20460

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                              Prepared by

                           Dr. Georgi Kazachki
                             CRYOTHERM
                1442 Wembley Ct. NE Atlanta, GA 30329
                                 Disclaimer
The views expressed in this report are those of the author and do not necessarily reflect those of EPA.
     Any mention of trade names or commercial products does not constitute endorsement or
                           recommendation for use.

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TABLE OF CONTENTS

1.  Introduction	1
2.  Study Approach	3
3.  Parameters affecting the performance and energy efficiency of a supermarket refrigeration
   system	6
4.  Design and operational features affecting the performance and energy efficiency of
   refrigeration systems	7
   4.1 Systems to be investigated	7
   4.2 Store size, location, and assumptions	7
   4.3 Conditions for the analysis	8
5.  Energy analysis methodology	16
   5.1 Number of bin hours	16
   5.2 System power input	16
       5.2.1 Power input into compressors	16
       5.2.2 Power input into circulation pumps	16
   5.3 Cooling load	21
   5.4 Bin energy consumption	24
   5.5 Annual energy consumption	24
6.  Results: bin and annual energy consumption of the baseline and alternative technologies ....25
7.  Analysis of the results	28
8.  Summary of conclusions and recommendations for next steps	31
   8.1 Summary of conclusions	31
   8.2 Recommendations for next steps	32

Appendices:

Appendix A:  Theoretical Analysis of Alternative Supermarket Refrigeration Technologies:
             Technical Review Committee Members	A-l
Appendix B:  EPA Supermarket Alternatives Study Report (August 6, 2007), Phase 1: Proposal
             for a detailed engineering analysis—description of a baseline store and alternative
             configurations	B-l
Appendix C:  Results tables: annual energy consumption, power input, and weather data, by bin
             and geographic location	C-l

List of Tables:

Table 1:   Conditions for the theoretical analysis	11
Table 2:   Descriptive conditions for the theoretical analysis	12
Table 3:   Weather bin data for Atlanta, GA; Boulder, CO;  and Philadelphia, PA	16
Table 4:   Performance table of a low-temperature compressor at return gas temperature 45°F
          and zero liquid-refrigerant subcooling	19


Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                       [ i ]


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Table 5:   Performance table of a medium-temperature compressor at return gas temperature
          45°F and zero liquid-refrigerant subcooling	20
Table 6:   Properties of inhibited Propylene Glycol 30% by weight, freezing point 9.2°F	21
Table?:   Properties ofDynaleneHC-30	21
Table 8:   Annual energy consumption of supermarket refrigeration technologies at three
          geographic locations	26
Table 9:   Number of hours of MT and LT compressors at their minimum operating SDT (50°F
          forMT and40°F forLT) at the three geographic locations	30
Table 10:  Conditions for a detailed engineering analysis 	34

List of Figures:

Figure 1:  Piping diagram of Baseline (DX) and Alternative C (DS)	13
Figure 2:  Piping diagram of Alternative A (MTS)	14
Figure 3:  Piping diagram of Alternative B (SC)	15
Figure 4:  HFC-404A pressure-enthalpy diagram with definitions of key parameters	18
Figure 5:  Annual energy consumption of supermarket refrigeration technologies in three
          geographic locations	27
[ii]
Final Report, September 2, 2008


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1.  INTRODUCTION

EPA is developing a voluntary partnership with the supermarket industry to facilitate the
transition from ozone-depleting substances to ozone-friendly alternatives. Known as the
GreenChill Advanced Refrigeration Partnership, the overall goal of this program is to promote
the  adoption of technologies, strategies, and practices that lower emissions of ozone-depleting
substances (ODS) and greenhouse gases (GHGs) through both the reduction of refrigerant
emissions and the increase of refrigeration systems' energy efficiency. Specific partnership goals
are  to provide supermarkets and organizations that support the supermarket industry with
information and assistance to:

    •   Transition to non-ODS refrigerants
    •   Reduce both ODS and non-ODS refrigerant emissions
    •   Promote supermarkets' adoption of alternative refrigeration technologies that offer
       qualities such as:
           o  Reduced ODS/GHG emissions (e.g., through reduced refrigerant charges and leak
             rates)
           o  Potential for improved energy efficiency
           o  Reduced maintenance and refrigerant costs
           o  Extended shelf life of perishable food products
           o  Improved system design, operations, and maintenance
    •   Reduce the total impact of supermarkets on ozone depletion and global warming

A key component of the GreenChill Partnership is to facilitate technological research and
information-sharing to assist partners in meeting these goals. EPA, in conjunction with the Food
Marketing Institute (FMI), determined that one area where information is currently limited
involves assessment of the energy efficiencies and energy consumption of currently available,
alternative supermarket refrigeration systems.  Consequently, EPA commissioned this study to
compare the energy consumption of alternative supermarket refrigeration technologies. The
study is based on theoretical analyses of the energy efficiency of the three most common
refrigeration technologies:

    •  Direct-expansion (DX) centralized systems. In a direct expansion system, the
       compressors of one suction group are mounted together on a rack and share suction and
       discharge refrigeration lines. Liquid and suction lines run throughout the store, feeding
       refrigerant to the cases and coolers and returning refrigerant vapor to the suction
       manifold. The compressor racks are located in a separate machine room, either in the
       back of the store inside or outside of the building, or on its roof, to reduce noise and
       prevent customer access. Condensers are usually air-cooled and hence are placed outside
       to reject heat. These multiple compressor racks operate at various suction pressures to
       support refrigerated fixtures (i.e., display cases, coolers, freezers, and some other small
       consumers) operating at different evaporating temperatures. The hot gas from the
       compressors is piped to the condenser and converted to liquid. The liquid refrigerant is
       then piped to the receiver and distributed to the fixtures by the liquid supply lines. After
       evaporating in the fixtures, the refrigerant returns in suction lines to the suction manifold
       and the compressors.

Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                       [ 1 ]


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    •  Secondary-loop, secondary-coolant, centralized systems (SC). Two fluids are used in
       secondary loop systems: the first is a secondary coolant, which is pumped throughout the
       store to remove heat from the refrigerated fixtures, and the second fluid is a refrigerant
       used to cool the cold fluid. Secondary loop systems can operate with two to four separate
       loops and chiller systems depending on the temperatures needed for the display cases.
       Secondary loop systems use a much smaller refrigerant charge than traditional direct
       expansion refrigeration systems.
    •  Distributed systems (DS).  Unlike traditional direct expansion refrigeration systems,
       which have a central refrigeration room containing multiple compressor racks,
       distributed systems use multiple smaller rooftop units that connect to cases and coolers,
       using considerably less piping. The compressors in a distributed system are located near
       the display cases they serve - on the roof above the cases, behind a nearby wall, or even
       on top of or next to the case in the sales area. Thus, distributed systems typically use a
       smaller refrigerant charge than DX systems. *

The analysis uses primarily existing thermo-physical data for refrigerants and secondary-coolant
fluids, as well  as performance characteristics from existing laboratory and field measurements,
and manufacturers' data. A significant attempt was made to reach beyond traditional
theoretical/academic studies and to reflect current best practices of the supermarket industry.
1 GreenChill Advanced Refrigeration Partnership Web Site. Advanced Refrigeration Technology.
hitp://www. cpa.gov/grccncMll/alitcchnologY. html
[ 2 ]
                                                              Final Report, September 2, 2008


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2.  STUDY APPROACH

This study was conducted with input from EPA and a Technical Review Committee, convened
by EPA, that includes GreenChill partners and EPA representatives (see Appendix A for a list of
GreenChill Technical Review Committee members). Cryotherm developed a study work plan
that outlined an approach for conducting a theoretical study comparing the energy usage for six
supermarket refrigeration scenarios. EPA and Cryotherm presented the initial work plan to the
Technical Review Committee during a conference call held on July 13, 2007, with a follow-up
call on August 9, 2007. Based on these discussions, the baseline and alternative scenarios were
redefined, three cities were chosen to represent different climates to be investigated, and a
detailed set of parameters that could affect the performance of supermarket refrigeration systems
was developed. Cryotherm and EPA presented initial results and conclusions of the theoretical
study at FMFs Energy and Technical Services conference held September 9-12, 2007 in Denver,
Colorado.

The general approach for conducting this study involved the following steps:

1.   Define Baseline and Alternative Scenarios.
    Based on input from the Technical Review Committee, the following baseline and
    alternatives were defined:

    Baseline:         New supermarket with a DX refrigeration system using an HFC
                    refrigerant (DX).

    Alternative A:    New supermarket with a low temperature (LT) DX and medium
                    temperature  (MT) glycol secondary loop refrigeration system using an
                    HFC refrigerant (MTS).

    Alternative B:    New supermarket with a LT secondary loop refrigeration system and a
                    MT secondary loop refrigeration system, each using an HFC refrigerant
                    (SC).

    Alternative C:    New supermarket with a distributed refrigeration system using an HFC
                    refrigerant (DS).

2   Identify geographic locations for study analysis.
    The Technical Review Committee, EPA, and Cryotherm selected three cities on which to
    conduct the analysis: Atlanta, Georgia; Boulder, Colorado; and Philadelphia, Pennsylvania.
    These cities were selected to represent both different climates in the U.S. and locations that
    are near the GreenChill partners' stores.

3.   Identify general parameters affecting the performance and energy efficiency of
    supermarket refrigeration systems (Section 3).
    Cryotherm developed a list of parameters affecting alternative supermarket refrigeration
    systems, based on a literature review and experience in designing and analyzing advanced
    refrigeration systems. Three groups of parameters were identified:

    •   Parameters determined by the ambient conditions at the location of the store,


Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                      [ 3 ]


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   •   Parameters determined by the indoor conditions in the store, and

   •   Parameters defined by the type of refrigeration system, its design features, and its
       interaction with the outdoor and indoor ambient conditions.

4.  Specify the design and operational features of each refrigeration system (Section 4 and
   Appendix B).
   This step involved considerable input from the Technical Review Committee. Based on an
   existing store layout (including piping and refrigerated fixtures, such as display cases,
   coolers, and freezers), the specific design and operational features also reflect the variety of
   technical and design approaches, geographic locations, store sizes,  and other experiences
   represented by the committee members and their supermarket chains. The list of parameters
   developed through this consensus process with the Technical Review Committee was
   presented in a Phase 1 report submitted to EPA on August 6, 2007  and is provided in
   Appendix B.

   The level of detail described for these parameters was appropriate for a detailed engineering
   analysis of the baseline and alternative scenarios. It was, therefore, necessary to use these
   specifications as the basis for defining a more simplified set of parameters that realistically
   reflect currently-designed supermarket refrigeration systems that could be analyzed from a
   more theoretical perspective. The temperature levels and refrigeration loads are based on
   actual store(s) recently or soon to be constructed. While the detailed set of parameters
   defined multiple temperature levels for the baseline and each alternative, the theoretical study
   assesses  a single temperature level  for the medium and low-temperature refrigeration systems
   (i.e., the Baseline and Alternatives  A and B)  and two temperature levels for the medium-
   temperature distributed system (i.e., Alternative C). The medium-temperature and the low-
   temperature refrigeration loads in the theoretical study are similar to the corresponding loads
   defined for a detailed engineering analysis. The  key conditions assumed for the theoretical
   study are described below and a more detailed description of these  parameters is provided in
   Section 4.

   •   Baseline: DX system consisting of a medium-temperature suction group with a saturation
       suction temperature at +20°F corresponding to evaporating temperature at the MT
       refrigerated fixtures at +22°F and a low-temperature suction group with a saturation
       suction temperature at -20°F corresponding to evaporating temperature at the LT
       refrigerated fixtures -18°F.

   •   Alternative A: Secondary-coolant medium-temperature system  with SST = 17°F
       providing +22°F supply temperature of the secondary coolant and a DX low-temperature
       suction group with a saturation suction temperature at -20°F.

   •   Alternative B: Secondary-coolant system consisting of a medium-temperature circuit
       with  a secondary-coolant supply temperature at +22°F and a low-temperature circuit with
       a secondary-coolant supply temperature at -18°F.

   •   Alternative C: Distributed  system consisting of two medium-temperature suction groups,
       at 20°F and 25°F, and one low-temperature suction group at -20°F.
[ 4 ]                                                          Final Report, September 2, 2008


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5  Develop Energy Analysis Methodology (Section 5).
   Cryotherm developed a methodology for estimating the annual energy consumption for the
   baseline and alternative scenarios at each of the geographic locations (i.e., Atlanta, Boulder,
   and Philadelphia). This involved estimating the power input into compressors and circulation
   pumps for each refrigeration circuit/system in the store differentiated by suction groups and
   supply temperatures of secondary coolants. The power input for a given suction group was
   determined as a function of the ambient temperature and the cooling capacity. The ambient
   temperatures were divided into 5°R groups (bins). The power input of each system/circuit
   was determined for the average temperature in each bin. The "WYEC2 Weather Year for
   Energy Calculations 2" software of the American Society of Heating, Refrigerating, and Air-
   Conditioning Engineers (ASHRAE) was used as a source of the weather data and hourly
   frequency of occurrence of each ambient temperature for the three locations. Manufacturers'
   data were used as the source of compressor performance data and system energy efficiency
   ratios (EER), used in calculating the compressors' power input. The EER was determined as
   a function of the saturated suction temperature (SST) in the analyzed system/circuit, useful
   superheat in the refrigerated fixture, return-gas temperature to the compressors, liquid-
   refrigerant subcooling into the refrigerated fixture, and the saturation discharge temperature
   (SDT). The saturated discharge temperatures were approximated with condensing
   temperatures. The condensing temperatures were correlated with the ambient temperature in
   each bin by adding a temperature difference of 10°R. In the range of ambient temperatures
   for which the compressor SDT would fall below the  set minimum, the EER at the minimum
   allowable SDT was used. The energy consumption at the average ambient temperatures in
   each bin was determined as the product of the corresponding power input in the bin and the
   number of hours in the bin for each location. The annual energy consumption was then
   calculated as the sum of the energy consumption in all bins.

6.  Conduct Analysis and Present Results (Section 6).
   Cryotherm conducted the energy analysis and described the study findings. The results
   compare energy consumption by type of system, by baseline vs. alternatives, and by location.
   For the baseline and each alternative, Cryotherm developed a set of three tables showing the
   energy consumption per bin and annual energy consumption at each location: Atlanta, GA;
   Boulder, CO, and Philadelphia, PA. Cryotherm summarized the results in a table by suction
   groups, technologies, and locations. The summary results are also presented graphically in a
   bar chart showing the annual energy consumption for each of the analyzed technologies.

7  Analyze Results (Section 7).
   Cryotherm analyzed the annual energy consumption results, comparing the energy
   consumption of each alternative with the baseline system, by geographical location. Factors
   that affect the energy efficiency and energy consumption of each alternative are discussed.

8  Present Conclusions and Recommendations for Next Steps (Section 8).
   This section presents final conclusions and suggests  next steps for future and/or more
   detailed analyses of the energy efficiency and energy consumption of alternative supermarket
   refrigeration systems.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                      [ 5 ]


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3. PARAMETERS AFFECTING THE PERFORMANCE AND
ENERGY EFFICIENCY OF A SUPERMARKET REFRIGERATION
SYSTEM
The major parameters affecting the performance and energy efficiency of a supermarket
refrigeration system reflect the ambient conditions, indoor conditions, and system design
features. The system operational parameters are a consequence of the system interaction with the
ambient and indoor conditions. The general parameters under consideration are:

   •  Ambient conditions
         o   Store location
         o   Ambient temperature

   •  Indoor data
         o   Indoor temperature
         o   Humidity

   •  System Design Features
         o   Refrigeration loads
         o   Suction saturation temperature
         o   Discharge saturation temperature
         o   Liquid refrigerant subcooling
         o   Refrigerant vapor superheat
         o   Type of system (e.g., DX, SC or DS)
         o   Refrigerant selection
         o   Secondary coolant selection
         o   Components selection
[6]
Final Report, September 2, 2008


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4.  DESIGN AND OPERATIONAL  FEATURES AFFECTING  THE
PERFORMANCE AND ENERGY EFFICIENCY OF
REFRIGERATION SYSTEMS

The theoretical study was performed based on the parameters, assumptions, and conditions that
affect refrigeration system performance and energy, described below. As described in the study
approach, these study parameters, assumptions, and conditions were developed based on input,
experience, and review from EPA and the Technical Review Committee. A summary of the key
conditions is presented in Table 1, and Table 2 describes the parameters organized by the
baseline and each alternative. Figures 1, 2, and 3 illustrate the piping layout for the DX baseline
and Alternative C (DS) systems, Alternative A (MTS), and Alternative B (SC), respectively.
4.1  Systems to be investigated
   Baseline:
   Alternative A:
   Alternative B:
   Alternative C:
Supermarket with a DX refrigeration system with HFC-404A as the
refrigerant (DX).

Supermarket with a low-temperature DX and medium-temperature
propylene glycol secondary-coolant refrigeration system using HFC-404A
as the refrigerant (MTS).

Supermarket with both MT and LT secondary-coolant refrigeration
systems using HFC-404A as the refrigerant in the primary systems (SC).

Supermarket with distributed refrigeration systems with HFC-404A as the
refrigerant (DS)
4.2 Store size, location, and assumptions

1.  The baseline and alternative stores are each 45,000 sq. ft.

2.  Stores consist of a medium-temperature (MT) refrigeration system with a refrigerating load
   of 856,079 Btu/h at a saturated suction temperature of+20°F and a low-temperature system
   (LT) with a refrigerating load of 300,000 Btu/h at a saturated suction temperature of-20°F.
   These loads were chosen to closely match the total load and approximate distribution in an
   actual store (i.e., recently or soon to be constructed).

3.  The refrigeration loads are from the refrigerated fixtures only. The load from the mechanical
   subcooling of the LT liquid refrigerant is added to the MT load.

4.  All systems use HFC-404A as the refrigerant.

5.  Locations for the analysis  are Atlanta, GA; Boulder, CO; and Philadelphia, PA.

6.  Heat reclaim and defrost method are excluded from the analysis.

7.  Heating and air-conditioning loads, building fire  and safety code, store lighting, plug loads
   and other loads, and the HVAC annual consumption are excluded from this study.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
                                                               7]


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                                                              Nomenclature

                                                DX    Direct expansion
                                                IHX   I nte rmed iate h e at exch anger
                                                      (e va porator/ch i lie r)
                                                LT    Low-te mpe ratu re
                                                MT    Medium-temperature
                                                MTS  Medium-temperature secondary
                                                MSC  Mechanical subcooling, °R
                                                NSC  Natural subcooling, °R
                                                RGT  Return-gas temperature, °F
                                                SC    Secondarycoolant
                                                SCSI Secondary-coolant supply temperature, °F
                                                SDT  Saturation discharge temperature, °F
                                                SST  Saturation suction temperature, °F
                                                TD    Temperature difference, °R
8.  The analysis for the baseline and all
   alternatives use the energy efficiency ratio
   (EER) of a representative compressor based
   on manufacturer's data calculated at the
   specified operating conditions for each
   alternative technology.

4.3. Conditions for the analysis

1.  Number of suction groups, secondary-coolant
   circuits and refrigeration loads:

   a.  Baseline: one LT DX suction group with a
       saturation suction temperature of -20°F,
       yielding an evaporating temperature of
       -18°F at the refrigerated fixtures and one
       MT DX suction group with a saturation
       suction temperature of +20°F, yielding  an
       evaporating temperature of 22°F at the
       refrigerated fixture.

   b.  Alternative A: one LT DX suction group with a saturation suction temperature of -20°F,
       yielding an evaporating temperature of-18°F, and one secondary-coolant circuit with
       SST 17 yielding a +22°F secondary-coolant supply temperature. The refrigeration loads
       from the refrigerated fixtures in the MT and LT circuits are the same as in the baseline.

   c.  Alternative B: one MT and one LT SC  circuit with +22°F and -18°F secondary-coolant
       supply temperature, respectively. The corresponding SST are 17°F in the MT and -23°F
       in the LT circuit. The refrigeration loads from the refrigerated fixtures  in the MT and LT
       circuits are the same as in the baseline.

   d.  Alternative C: one LT DX suction group with saturation suction temperature -20°F and
       two MT suction groups with saturation  suction temperatures of+25°F  and +20°F. The LT
       refrigeration load from the refrigerated  fixtures is the same as in the baseline. The MT
       refrigeration load is distributed as follows: 450,000 Btu/hr at 20°F and 406,079 Btu/hr at
       SST at 25°F. The load from the mechanical subcooling of the LT liquid refrigerant is
       added to the load of the group with SST of 25°F.

2.  Compressor return gas temperature: 45°F

3.  Useful superheat in the DX refrigerated fixtures, mechanical sub-cooler, and intermediate
   heat exchanger (IHX):

   a.  MT: 5°R

   b.  LT: 15°R

   c.  Mechanical sub-cooler and IHX: 10°F

4.  Mechanical subcooling (MSC) of the LT liquid refrigerant by the MT refrigerant:

   a.  Baseline: to 50°F
[8]
                                                            Final Report, September 2, 2008


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   b.  Alternative A (MTS): to 50°F.
   c.  Alternative B (SC): to 50°F, 40°F, and 30°F.
   d.  Alternative C (DS): to 50°F.
5.  Impact of heat gains/losses in the liquid refrigerant lines on subcooling at the display cases
   and intermediate heat exchanger (IHX): neglected.
6.  Heat gains in DX return lines and in SC supply and return lines: neglected.
7.  Condenser temperature difference: 10°R for both MT and LT in all technologies.
8.  Natural subcooling in the condensers: 0°R for all systems.
9.  Condenser fan control:
   a.  Baseline (DX): float SDT to 70°F for MT and LT condensers.
   b.  Alternative A (MTS): float SDT to 50°F for MT and to 70°F for LT condensers.
   c.  Alternative B (SC): float SDT to 50°F for MT and to 40°F for LT condensers.
   d.  Alternative C (DS): float SDT to 70°F for both MT and LT condensers.
   While some supermarket DX systems operate at 50°F SDT, this study assumes floating the
   condensing temperature to 70°F for the DX systems and 50°F or 40°F for the SC systems.
   This accounts for the long refrigerant lines in DX systems and the possibility of the liquid
   refrigerant reaching saturation point at the expansion valves, resulting in malfunction. The
   shorter liquid refrigerant lines in an SC system allow floating the condensing temperature to
   lower temperatures without causing problems at the expansion valves.
   The MT SST in the Baseline DX and in Alternative C is assumed to be 20°F. Accounting for
   a 2°R equivalent pressure drop in the suction line for oil return, this yields an evaporating
   temperature of 22°F in the evaporator. The MT SST in Alternative A and Alternative B are
   17°F. The LT SST in the Baseline DX and in Alternative C is assumed to be -20°F yielding
   an evaporating temperature of-18°F in the evaporator.
   Assuming a 5°R temperature difference in the MT IHX, a SST of 17°F yields 22°F
   secondary fluid going to the refrigerated fixtures.  Thus, the MT SST in Alt A and B are 3°R
   lower than the corresponding MT SST in the Baseline DX. Similarly, the LT SST is -20°F
   for the baseline DX and -23°F for Alternative B.
10. Compressor inlet pressure:
   a.  Pressure drop in DX baseline, DX LT, and DS MT and LT suction lines: 2°R equivalent
       for oil return
   b.  Pressure drop in the Alternatives A (MTS) MT and B (SC) MT and LT suction lines:
       neglected because of the short return lines and the downstream movement of oil.
11. Compressor inlet temperature (Return Gas Temperature): 45°F in DX and DS,  10°R
   superheat in SC.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                       [ 9 ]


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12. Secondary-coolant supply/return temperature difference: 7°R
13. Circulation pumps:
   a.   The power input into the SC circulation pumps is added to the power input of the
       compressor racks.
   b.   90% of the heat from the pumps is added to the cooling load from the fixtures.
   c.   Pressure head of the LT and MT SC circulation pumps is assumed to be 70 ft.
   d.   Assumed efficiency (including electric motor efficiency) of the LT and MT SC
       circulation pumps is 60%.
14. Analysis with Dynalene in the LT SC and Propylene Glycol in the MT SC.
15. Indoor temperature and relative humidity for the study: 75°F 755% year around.
[10;
Final Report, September 2, 2008


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Table 1: Conditions for the theoretical analysis
Technology

Baseline

Alternative A

Alternative B

Alternative C


System
Type
DX
DX
DX
SC
SC
SC
DS
DS
DS
Temp.
Level
LT
MT
LT
MT
LT
MT
LT
MT
MT
Notes

Subcooled by MT

Same as Baseline

Subcooled by MT

Subcooled by MT


SST
°F
-20
20
-20
17
-23
17
-20
25
20
Max SDT
°F
110
110
110
110
110
110
110
110
110
Min SDT
°F
70
70
70
50
40
50
70
70
70
Liquid Temp.
°F
50
SDT
50
SDT
50, 40, 30
SDT
50
SDT
SDT
Refrigerant/Sec. Coolant Temp.
at Case/Chiller Outlet
°F
-5
25
-5
27
-13
27
-5
30
25
RGT
°F
45
45
45
27
-13
27
45
45
45
Cooling load
Btu/hr
300,000
856,079 + MSC
300,000
856,079 + MSC + PH
300,000 + PH
856,079 + MSC + PH
300,000
406,079 + MSC
450,000
Power
kW



addPP
addPP
addPP



MSC = Mechanical subcooling, PH = Pump heat, PP = Pump power, SDT = Saturation discharge temperature, RGT = Return gas temperature
In the theoretical analysis, 2°R of equivalent pressure drop for oil return in the LTDX and MTDX lines has been assumed.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
                       11]
www.epa.gov/gree n chill
Chiltin' for the environment

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Table 2: Descriptive conditions for the theoretical analysis
Analysis is based on a supermarket refrigeration system with cooling loads close to that in a real store.
Baseline DX system
One MT system 856,079 Btu/h designed for SST/SDT = +20/110°F
One LT system 300,000 Btu/h designed for SST/SDT = -20/110°F, subcooled by MT
Condenser TD = 10.0°R for both MT and LT
Floating condensing pressure to: 70°F in both MT and LT condensers.
Equivalent pressure drop in suction lines: 2°R in both MT and LT systems
Natural subcooling NSC=0°R in both MT and LT systems
Mechanical subcooling of LT  liquid refrigerant in MT system to 50°F
Return gas temperature 45°F in both MT and LT systems resulting from:
  • 25°R MT compressor superheat of which 5°R useful superheat in MT evaporators/display cases and 20°R estimated
superheat in the return lines
  • 65°R LT compressor superheat of which 15°R useful superheat in LT evaporators/display cases and 50°R estimated
superheat in the return lines
Alternative A, MTS System
One SC MT system 856,079  Btu/h designed for SCST = +22°F, SST/SDT = +17/110°F
One DX LT system 300,000 Btu/h designed for SST/SDT -20/110°F, subcooled by MT
Condenser TD = 10.0°R for both MT and LT
Floating condensing pressure to: 50°F in MT and 70°F in LT condensers.
Equivalent pressure drop in suction lines: 0°R in MT SC and 2°R in LT DX
Natural subcooling NSC=0°R in both MT and LT systems
Mechanical subcooling of LT  liquid refrigerant in MT system to 50°F
Return gas temperature +27°F in the MT refrigerating circuit resulting from:
  • 10°R compressor superheat of which 10°R useful superheat in the intermediate heat exchanger and 0°R superheat in
refrigerant return lines (assumed no heat gains because of short lengths.)
Return gas temperature +45°F in LT resulting from:
  • 65°R LT compressor superheat of which 15°R useful superheat in LT evaporators/display cases and 50°R estimated
superheat in the return lines
Alternative B, SC system
One SC MT system 856,079  Btu/h designed for SCST = +22°F, SST/SDT = +17/110°F
OneSC LT system 300,000 Btu/h designed for SCST/SDT-18/110°F, SST/SDT =-23/110°F, subcooled by MT
Condenser TD =10.0°R for both MT and LT
Floating condensing pressure to: 50°F in both MT and LT condensers.
Equivalent pressure drop in suction lines: 0°R in both MT and LT
Natural subcooling NSC=0°R
Mechanical subcooling of LT  liquid refrigerant in MT system to 50, 40, and 30°F
Return gas temperature +27°F in MT and -13°F in  LT resulting from: 10°R useful superheat in both MT and LT IHX
Pump Design  Head both in MT and LT:  70 ft. H2O
Evaporator Design Temp. Difference both in MT and LT: 7°R
LT Secondary Coolant: Dynalene HC-30
MT Secondary Coolant: 30%  Propylene Glycol
Pump Efficiency, both MT and LT: 0.6 (including electric motor efficiency)
Pump Heat (% of Pump Work) both in MT and LT:  90%
Alternative C, DS system
One MT system 450,000 Btu/h designed for SST/SDT = +20/110°F
One MT system 406,000 Btu/h designed for SST/SDT = +25/110°F
One LT system 300,000 Btu/h designed for SST/SDT = -20/110°F
Condenser TD = 10.0°R for both MT and LT
Floating condensing pressure to: 70°F in both MT and LT condensers.
Equivalent pressure drop in suction lines: 2°R in both MT and LT
Natural subcooling NSC=0°R
Mechanical subcooling of LT  liquid refrigerant in MT system with SST +25°F: to 50, 40, and 30°F
Return gas temperature 45°F in both MT and LT systems resulting from:
  • 25°R MT compressor superheat of which 5°R useful superheat in MT evaporators/display
  cases and 20°R estimated  superheat in the return lines
  • 65°R LT compressor superheat of which 15°R useful superheat in LT evaporators/display
  cases and 50°R estimated  superheat in the return lines
Summary of the general assumptions:
Refrigeration load from the refrigerated fixtures is independent of operating conditions (except for LT subcooling load)
Condenser TD is 10.0°R
DX Systems designed with 2°R equivalent pressure drop in  suction lines, SC with 0°R
[12]                                                                 Final Report, September 2, 2008


-------
 Figure 1: Piping diagram of Baseline (DX) and Alternative C (DS)
   BASELINE SYSTEM, ALTERNATE C (DISTRIBUTED)

          MEDIUM-TEMPERATURE SYSTEMS
                                                  LOW-TEMPERATURE SYSTEMS
                                      COMPONENTS AT REFRIGERATION SYSTEM
                   *
•9
                                                                  Q
                                                                    CONDENSER
VCOND.
                                                                      CONDENSER

                                                                       WIT     COUP.
                                                                       COND.
                                                                              W
                                                                     LTCOMP.
                                                     MECHANICAL
                                                     SUBCOOLER
                                                                 QSUBCOOL
  \'<£a|
                       . LINE

                                           DISTRIBUTION PIPING
                                        Q
                                RET. LINE
                                                                         . LINE
     EXPANSION
     DEVICE  '
                   EVAPORATOR COIL
            '  >   [5yy
               Q
                                         COMPONENTS IN DISPLAY-CASE
                                            OR UNIT-COOLER
                 MT LOADS, SUBCOOL
                                                            LT LOADS
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
www.epa.gov/greenchill
                                                                                        13]
                                                                 Chiltin' for the environment

-------
Figure 2: Piping diagram of Alternative A (MTS)



     ALTERNATE A (MTSC, LTDX)



            MEDIUM-TEMPERATURE SYSTEMS
                                                         LOW-TEMPERATURE SYSTEMS
[14;
www.epa.gov/greenchill
                                                                Final Report, September 2, 2008
                                                                     Chiltin' for the environment

-------
 Figure 3: Piping diagram of Alternative B (SC)


        ALTERNATE B (MTSC, LTSC)


               MEDIUM-TEMPERATURE SYSTEMS
LOW-TEMPERATURE SYSTEMS
                   EXPANSION  MECHANICAL
                    DEVICE ^ SUBCOOLER

                                o
                                         COMPONENTS AT REFRIGERATION SYSTEM
                                            MT COMP.
                                            MT PUMP
                            LT COMP.
                                                DISTRIBUTION PIPING
                                             RET. LINE
                                                                   Q
      SUP. LINE
                                              COMPONENTS IN DISPLAY-CASE
                                                 OR UNIT-COOLER
                                                                      HEAT EXCHANGER
                                                                 ->"-towyAoj-^—'

                                                                         ^Q
                            LTPUMP
RET. LINE
                                                                            
-------
5.  ENERGY ANALYSIS METHODOLOGY
The energy analysis and comparison of the three alternative technologies with the baseline DX
technology was performed based on an estimation of the annual energy consumption at three
geographic locations. The annual energy consumption of the baseline and the alternative
technologies was determined from the power input into the MT and LT refrigeration systems and
the  number of operating hours. Since both of these factors vary with the ambient temperature, the
calculation was performed across the range of ambient temperatures for each of the three
geographic locations during a year. For practical purposes, the range of ambient temperatures
was divided into temperature intervals, or "bins."

5.1. Number of bin hours

Weather statistical data for the three analyzed geographic locations provided the number of hours
the  temperatures in each bin occur in a year. The source of these weather data was ASHRAE' s
WYEC2 Weather Year for Energy Calculations 2. The bin hours for the three locations analyzed
in this study are shown in Table 3.
Table 3: Weather Bin Data for Atlanta, GA; Boulder, CO; and Philadelphia, PA
Ambient Temperature Weather Bin Data Weather Bin Data Weather Bin Data
Bin Atlanta, GA Boulder, CO Philadelphia, PA
°F Hours Hours Hours
95-100 9 22 3
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Total Hours
56
196
758
768
1314
885
1027
790
673
641
436
560
323
181
72
64
7
0
0
8760
96
115
382
440
489
503
907
698
754
762
633
834
717
611
251
201
130
89
126a
8760
52
104
477
656
907
619
983
625
540
576
552
1067
685
442
248
184
40
0
0
8760
3 The number of hours in this bin is the cumulative number of hours of all temperatures within and below the 5°F-
bin, thus integrating the 5°F bin and the next lower-temperature bins. The reason is that all these temperatures
affect the performance of the refrigeration system in a similar way and can be processed together.
                                                           Final Report, September 2, 2008


-------
5.2. System power input
The power input into the refrigeration system consists of the power input into the refrigerating
compressors, condenser fans, secondary-coolant circulation pumps, refrigerated fixture lights and
fans, anti-sweat heaters, and defrost heaters. This study assessed only the power input into
compressors and circulation pumps. The power input into refrigerated fixture lights and fans,
anti-sweat and defrost heaters were omitted since they affect all technologies equally. In this
study, the power input into condenser fans is assumed to be equal among refrigeration
technologies; however, in reality there are slight differences. An exact engineering analysis could
account for theses differences.
When determining the system power input, it was assumed that it depends only on the ambient
temperature and not on the specific time when this ambient temperature occurs. Thus, the energy
consumption in each bin reflects the number of hours in the bin and the system power input at
the average temperature in the bin. The annual energy consumption is a sum of the energy
consumption of all bins.
5.2.1 Power input into compressors
The compressor power input was determined from the system cooling load and the net EER from
equation (1). The system cooling load is discussed in Section 5.3. The net EER is determined
from the compressor performance characteristics by the SST, SDT, return gas temperature, liquid
subcooling, and useful superheat. (These parameters are specified in detail for each technology
in Table 2 and illustrated in Figure 4.) The condensing temperature in all technologies was
determined by adding the specified temperature difference of 10°R to the ambient temperature.
This applies in the range from the highest to the lowest  ambient temperature bins at which the
condensing temperature has reached its specified minimum value, designated as lowest floating
condensing pressure/temperature (see Table 2).
Compressor manufacturers provide compressor performance data in a variety of formats,
including tables, curves, equations, and software packages. For this study, performance data for
Copeland brand compressors were used. The performance data for LT compressors were derived
from the compressor 3DRHF46KE-TFC and are shown in Table 4. and the performance data for
MT compressors were derived from the compressor 3DS3R17ML-TFC and are shown in Table
5.
The power input at the average ambient temperature in each bin was determined from the system
cooling capacity and the net EER by applying the following equation:
                                   System cooling capacity [Btu/h]
   Compressor Power Input [W] =	     (1)
                                 Net Energy Efficiency Ratio [Btu/Wh]
5.2.2 Power Input into Circulation Pumps
The power input into the secondary-coolant circulation  pumps was determined from the
following equation:
          Volumetric Flow Rate [m3/s]. Pressure difference [Pa]
   Pc.rc.pump [W] =	       (2)
                              Efficiency

Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                     [ 17 ]


-------
 Figure 4: HFC-404A pressure-enthalpy diagram with definitions of key parameters
    .2
    a.

    2
    3
    I/)
    I/)
    O
        /in 1111/11111 i/i 11111 i/i 111111/111111 n 111111/11
      75.0  80.0  850  900  95.0  100.  105   110  115   120  125   130.  135.  HO.  US.  150  155  1
                                            Enthalpy, Btu/Lb
 Figure definitions of parameters used in calculating the performance of the refrigeration system and
 refrigerating compressors: saturation suction temperature (SST), saturation discharge temperature (SDT),
 subcooling, useful superheat, non-useful superheat (in the return lines), and return gas temperature.
 Compressor performance characteristics refer to compressor superheat which is the sum of the non-useful
 and useful superheat. Non-useful superheat is used here only for illustration purposes. In the first order of
 simplification, suction saturation temperature is used interchangeably with evaporating temperature and
 saturation discharge temperature is used interchangeably with condensing temperature. In this picture:
 SST = 20°F, SDT = 110°F, subcooling = 0°R, useful superheat = 5°R, compressor superheat = 25°R, and
 return gas temperature = 45°F. The source of refrigerant properties is  NIST Refprop 7.0.
[18;
Final Report, September 2, 2008
www.epa.gov/greenchill
     Chiltin' for the environment

-------
Table 4: Performance table of a low-temperature compressor at return gas temperature 45°F and
zero liquid-refrigerant subcooling. C = Capacity, Btu/hr; P = Power, W; A = Current, A; M = Refrigerant
mass flow rate, Ib/hr; E = EER, Btu/W-hr; % = Isentropic Efficiency, %. Afofe: This table is used as an
illustration by permission from Emerson Climate Technologies.

CRATING CONDITIONS ^
45° F Return Gas
0°F Subcooling
95°F Ambient Air Over
v J
60 h
Conden
(Sat Dew
130
(354) C
P
A
IV
E
%
120 C
(310) P
IV
E
%
110 C
(271) P
IV
E
%
105 C
(252) P
IV
E
%
100 C
(235) P
IV
E
%
80 C
(174) P
IV
E
%
70 C
(148) P
IV
E
%
50 C
(104) P
IV
E
%
40 C
(86) P
IV
E
%
z Opera
sing Tern
Pt Press
-40(4.5)
17900
6350
21.9
432
2.8
64.7
22700
6550
22.3
495
3.5
68.9
26600
6600
22.4
530
4
69.8
28200
6550
22.3
545
4.3
69.6
29700
6550
22.2
550
4.5
69.1
34500
6200
21.6
560
5.6
65.9
36400
5900
21.2
560
6.2
64.2
40200
5200
20.5
555
7.8
61.5
42400
4790
20.4
560
8.8
60.5
tion
perature
ure, psig
-35(7.1)
2250C
7250
23.9
545
3.1
67.8
2740C
7350
24.1
60C
3.7
70.2
3140C
7300
23.9
63C
4.3
70.4
3310C
7250
23.8
64C
4.6
70
3480C
7150
23.6
645
4.8
69.5
401 OC
6650
22.6
655
6
66.6
4250C
6350
22.1
655
6.7
65.1
4730C
5550
21.2
655
8.5
62.6
5000C
5100
21
665
9.8
61.5
"F
) Evar.
-30(9.9)
2730C
8150
25.9
665
3.4
69.5
3230C
8100
25.9
71C
A
70.8
3660C
8000
25.6
735
4.6
70.7
3850C
7900
25.3
745
4.9
70.2
4030C
7800
25
75C
5.2
69.7
4640C
7150
23.7
76C
6.5
67
4930C
6750
23
76C
7.3
65.7
5500C
5900
21.9
77C
9.4
63.2
5850C
5450
21.6
78C
10.8
61.9
TE1V
orating 1
-25(13)
3230C
900C
28
79C
3.6
70.5
37600
8900
27.7
830
4.2
71.1
42200
8700
27.2
850
4.8
70.7
44300
8600
26.8
860
5.2
70.3
46300
8450
26.5
865
5.5
69.8
53500
7650
24.8
875
7
67.3
57000
7250
24
880
7.9
66.1
64000
6250
22.7
895
10.2
63.4
68500
5800
22.3
910
11.8
61.6
LO1
1PER
emperati
-20(16)
37600
9850
30
925
3.8
71
43200
9650
29.5
955
4.5
71.1
48300
9400
28.8
980
5.1
70.6
50500
9250
28.4
985
5.5
70.2
53000
9050
27.9
990
5.8
69.7
61000
8150
26
1000
7.5
67.5
65500
7700
25
1010
8.5
66.2
74000
6650
23.5
1040
11.2
63.1
79500
6150
23.1
1060
12.9
60.8
W
ATUI
jre °F (S<
-15(20)
43200
10700
32
1060
4
71.2
49300
10400
31.3
1100
4.7
71.1
55000
10100
30.4
1120
5.4
70.5
57500
9850
29.9
1120
5.8
70.1
60000
9650
29.4
1130
6.2
69.6
69500
8650
27.2
1150
8
67.4
74500
8150
26.1
1160
9.2
66.1
85500
7050
24.4
1200
12.1
62.3

*E
at Dew Pt
-10(24)
49100
11500
34
1220
4.3
71.2
56000
11200
33.1
1250
5
70.9
62000
10700
32
1270
5.8
70.3
65000
10500
31.5
1270
6.2
69.9
68000
10200
30.9
1280
6.6
69.5
79000
9150
28.4
1310
8.6
67.2
85000
8600
27.2
1320
9.9
65.7
97500
7450
25.3
1370
13.1
61

^3DRHF46KE-TFC ^
COPELAMETIC® HFC^04A
DISCUS® COMPRESSOR
TFC 208/230-3-60
V J
Pressure
-5(28)
55500
12300
36
1380
4.5
71.1
6300C
11900
34.9
141C
5.3
70.7
69500
11400
33.7
143C
6.1
70.1
7300C
11100
33
144C
6.6
69.7
7650C
10800
32.3
145C
7.1
69.2
8950C
9650
29.6
148C
9.3
66.7
9600C
9000
28.3
151C
10.7
65
111000
7800
26.3
1560
14.2
59.2

!, psig)
0(33)
62500
13000
38
1560
4.8
71
7050C
12500
36.7
1590
5.6
70.5
7800C
12000
35.3
1610
6.5
69.8
8200C
11700
34.6
1620
7
69.4
8550C
11400
33.8
1630
7.5
68.9
10100C
10100
30.8
1680
10
66.1
10800C
9450
29.5
1700
11.5
63.9
125000
8200
27.3
1780
15.3
56.9



NON-STANDARD CONDITIONS: Nominal Performance Values (±10%) based on 72 hours run-in. Subject to change without notice. Current @ 230 V
C:Capacity(Btu/hr|, P: Power) Watts), A:Current(Amps|, MiMass Flow(lbs/hr|, E:EER(Btu/Watt-hr|, %:lsentropic Efficiency(%|
©2007 Emerson Climate Technologies, Inc. - 1.24LD60-06-334-TFC
Autogenerated Compressor Performance ^ Printed 11/11/2007
06-334
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
19]


-------
Table 5: Performance table of a medium-temperature compressor at return gas temperature 45°F
and zero liquid-refrigerant subcooling. C = Capacity, Btu/hr; P = Power, W; A = Current, A; Wl =
Refrigerant mass flow rate, Ib/hr; E = EER, Btu/W-hr; % = Isentropic Efficiency, %. Note: This table is used
as an illustration by permission from Emerson Climate Technologies.
C RATING CONDITIONS
45°F Return Gas
0°F Subcooling
95°F Ambient Air Ove
V
60 h
Conden
(Sat Dew
140
(402) C
P
A
M
E
°/
130 C
(354) P
M
E
01
120 C
(310) P
A
M
E
%
110 C
(271) P
M
E
%
100 C
(235) P
M
E
01
90 C
(203) P
A
M
E
01
70 C
(148) P
M
E
O/
60 C
(125) P
M
E
%
50 C
(104) P
M
E
01
Iz Opera
sing Tern
Pt Press
-10(24)
42300
12000
38.8
1190
3.5
70.5
49000
11600
38
1220
4.2
70.2
55500
11200
37.2
1240
4.9
70
62000
10800
36.2
1270
5.7
69.7
68500
10400
35.1
1290
6.6
69.4
75000
9900
34.1
1320
7.6
68.8
89000
8800
32
1390
10.1
67.1
96000
8200
31.1
1420
11.7
65.7
104000
7550
30.3
1460
13.7
63.8
tion
perature
ure, psig
0(33)
54500
13600
42. e
1560
A
72.2
6300C
13100
41 .e
1580
4.8
71.3
71000
12600
40.4
1610
s.e
70.7
79000
12100
39.1
1640
e.e
70.2
8750C
11500
37.7
1670
7.e
69.5
9550C
10900
36.3
1700
8.8
68.7
113000
9550
33.e
1780
11.9
66.1
122000
8800
32.3
1820
13.9
64.1
13200C
8000
31.1
1870
16.5
61.2
MEDIUM
TEMPERATURE
	 /

°F
) Evaf
5(38)
61500
14400
44.6
1770
4.3
72.3
7050C
13900
43.4
1790
5.1
71.4
79500
13300
42
1810
e
70.7
89000
12700
40.6
1850
7
70.1
9800C
12000
39
1880
8.2
69.4
1 08000
11400
37.4
1920
9.5
68.5
1 27000
9850
34.2
2010
12.9
65.6
138000
9000
32.7
2060
15.3
63.2
14800C
8150
31.3
2110
18.2
59.7
MFCs Require Use of Polyol Ester
Lubricant Approved by Bulletin
AE-1248
>orating "I
10(44)
68500
15200
46.6
1990
4.5
72
7900C
14600
45.2
2010
5.4
71.2
89000
14000
43.7
2040
6.4
70.5
99500
13300
42
2080
7.5
70
11000C
12600
40.2
2130
8.8
69.3
12100C
11800
38.4
2170
10.3
68.3
1 43000
10100
34.8
2270
14.2
64.9
155000
9200
33.1
2330
16.9
62
16700C
8200
31.4
2380
20.3
57. £
emperati
15(49)
76000
16000
48.6
2230
4.7
71.6
8750C
15300
47
2260
5.7
70.8
99500
14600
45.3
2300
6.8
70.3
1 1 1 000
13900
43.4
2340
8
69.7
12300C
13100
41.4
2390
9.4
69
13500C
12200
39.4
2450
11.1
67.9
160000
10300
35.3
2560
15.6
63.9
1 73000
9300
33.3
2620
18.7
60.5
18700C
8200
31.4
2690
22.7
55.5
jre °F (S.
20(56)
84000
16800
50.6
2500
5
71.1
9750C
16100
48.8
2530
6
70.4
1 1 1 000
15300
46.9
2580
7.2
69.9
124000
14400
44.8
2630
8.6
69.3
13700C
13500
42.5
2690
10.2
68.5
15100C
12600
40.3
2750
12
67.3
1 79000
10500
35.6
2880
17.1
62.7
194000
9350
33.4
2950
20.8
58.6
20900C
8150
31.2
3030
25.6
52.6

at Dew Pt
25(63)
92500
17600
52.7
2800
5.3
70.5
108000
16800
50.6
2840
6.4
69.9
1 23000
15900
48.4
2890
7.7
69.4
1 38000
14900
46.1
2960
9.2
68.8
153000
13900
43.6
3020
11
67.9
1 68000
12900
41
3090
13.1
66.5
200000
10600
35.8
3240
18.9
61
21 7000
9350
33.3
3320
23.2
56.1
23400C
8050
30.8
3400
29.1
48.7
^3DS3R17ML-TFC ^
COPELAMETIC® HFC-404A
DISCUS® COMPRESSOR
TFC 208/230-3-60
V J
Pressure
30(70)
102000
18500
54.7
3130
5.5
69.8
119000
17500
52.4
3170
6.8
69.2
1 36000
16500
49.9
3240
8.2
68.8
1 53000
15400
47.3
3310
9.9
68.1
1 70000
14300
44.5
3390
11.9
67.1
1 87000
13200
41.7
3470
14.2
65.4
223000
10600
35.9
3630
21
58.9
241000
9250
33.1
3720
26.1
52.9
261 OOC
7850
30.3
3820
33.3
43.6
, psig)
35(78)
112000
19200
56.7
3500
5.8
69.1
131000
18200
54.2
3550
7.2
68.5
1 50000
17100
51.4
3630
8.8
68
1 69000
15900
48.5
3710
10.6
67.3
188000
14700
45.4
3800
12.8
66
208000
13400
42.3
3890
15.5
64
247000
10600
35.9
4070
23.3
56.1
268000
9100
32.7
4180
29.4
48.9

40(0)
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0
0


45(0)
0
0
0
0
0
0
c
0
0
0
0
c
0
0
0
0
0
c
0
0
0
0
0
0
c
0
0
0
0
0
c
0
0
0
0
c
0
0
0
0
0
c



NON-STANDARD CONDITIONS: Nominal Performance Values (±10%) based on 72 hours run-in. Subject to change without notice. Current® 230V
C:Capacity(Btu/hr), P:Power(Watts), A:Current(Amps), M:Mass Flow(lbs/hr , E:EER(Btu/Watt-hr), %:lsentropic Efficiency(%)
© 2007 Emerson Climate Technologies, Inc. .c 1.24MD60-06-5209-TFC
Autogenerated Compressor Performance T Printed 11/11/2007
06-5209
[20;
Final Report, September 2, 2008


-------
 The secondary-coolant volumetric flow-rates in both the MT and LT systems were determined
                               from the following equation:
               Cooling Capacity [Btu/hr]
   V [ft3/hr] =	
         Density [lb/ft3].Specific heat [Btu/lb-°R].Delta T [°R]
                                                      (3)
The density and the specific heat for propylene glycol, the secondary coolant in the MT system,
are shown in Table 6. The density and specific heat for Dynalene HC-30, the secondary coolant
in the LT system, are shown in Table 7. The pressure difference is derived from the pressure
head in the secondary-coolant systems. The pressure head in both MT and LT systems is 70 ft.
H2O. The delta T is the temperature difference between the secondary coolant supply and return
temperatures and is 7°R in both MT and LT  systems. The efficiency of the circulation pumps in
both MT and LT circuits is 60%.
 Table 6: Properties of inhibited Propylene
 Glycol 30% by weight, freezing point 9.2°F
  Fluid Temp.       Density       Specific Heat
      [°F]           [Ib/ft3]         [Btu/lb'R]
        10
        15
        20
        25
        30
        35
        40
        45
        50
        55
        60
        65
        70
64.96
64.91
64.86
64.81
64.75
64.69
64.63
64.57
64.50
64.43
64.36
64.28
64.21
0.901
0.902
0.904
0.906
0.908
0.910
0.911
0.913
0.915
0.917
0.919
0.921
0.922
Table 7: Properties of Dynalene HC-30
Temperature
°F
425
70
60
50
40
30
20
10
0
-10
-20
-30
Density
Ib/ft3
73.29
79.56
79.74
79.91
80.09
80.27
80.44
80.62
80.80
80.97
81.15
81.33
Specific Heat
Btu/lb°R
0.8447
0.7360
0.7329
0.7298
0.7268
0.7238
0.7206
0.7176
0.7145
0.7114
0.7084
0.7054
5.3. Cooling load
The major portion of the required cooling capacity used in the calculation of the compressor
power input is the cooling load from the display cases, coolers, and freezers. This cooling load is
often referred to as a net refrigerating load and is used to determine the system net refrigerating
capacity or net refrigerating effect. Additional cooling loads come from small local air-
conditioning units and from the mechanical subcooling of the LT liquid refrigerant in the MT
refrigeration system.
For efficient operation, the cooling loads are distributed into suction groups. Not all of the above
listed load components are present in each suction group. For instance, the mechanical
subcooling is piped into the MT with the highest SST. For this study, all MT net cooling loads in
the Baseline (DX system) have been combined into one suction group at SST of+20°F and  all
LT net loads have been combined into one suction group at SST -20°F. The combined MT net

Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                     [ 21 ]


-------
cooling load is 856,079 Btu/hr. The combined LT net cooling load is 300,000 Btu/hr. The load
from the mechanical subcooler is an additional load to the MT circuit.

The net cooling loads in Alternative A (MTS) have been serviced by one MT secondary-coolant
circuit with SCST of+22°F with an associated refrigerant SST of+17°F in the intermediate heat
exchanger (evaporator/chiller) and a net load of 856,079 Btu/hr.  The load from the mechanical
subcooler is added to this circuit. The heat gains from the SC circulation pumps are also added to
this circuit. Similar to the baseline, all LT net loads have been combined into one suction group
at SST -20°F. The combined LT net cooling load is 300,000 Btu/hr.

The cooling loads in Alternative B (SC) have been serviced by one MT secondary-coolant circuit
with SCST (secondary-coolant supply temperature) of+22°F with corresponding refrigerant SST
of+17°F and one LT secondary-coolant circuit with SCST -18°F with corresponding refrigerant
SST of-23°F. The MT circuit also includes the load from the mechanical subcooling of the LT
liquid refrigerant and the heat gains from the MT SC circulation pump. The LT circuit includes
the heat gains from the LT SC circulation pump. The net refrigeration loads from the fixtures are
the same as in the baseline system: MT 856,079 Btu/hr and LT 300,000 Btu/hr.

The MT refrigeration load in Alternative C (DS) is distributed between two suction groups: with
SST of+25°F and SST of+20°F to illustrate and assess the benefit of the distributed technology.
The net loads are 406,079 Btu/hr and 450,000 Btu/hr respectively. The first suction group also
assumes the load from the mechanical subcooling of the LT liquid refrigerant. Similar to the
baseline,  all LT net loads in Alternative C have been combined into one suction group at SST -
20°F. The combined LT net cooling load is 300,000 Btu/hr.

The net loads described above closely match the cooling loads in the supermarket store that was
selected as a reference store for this study. The analysis of combined MT and LT suction groups
in this study provides an objective tool for energy comparison of the alternative and baseline
technologies. A detailed engineering analysis would be required to assess a refrigeration system
with multiple suction/supply groups in all technologies.

The heat gains into the refrigerant return lines create additional load. In this study, the heat gains
into return lines are accounted for by an estimated vapor superheat between the outlet of the
display cases/evaporators and compressor inlet. This superheat is designated as a non-useful
superheat in Figure 4. A more detailed investigation of the impact  of the heat gains and other
parasitic losses in various parts and components of the baseline refrigeration system and
alternative technologies would require a more detailed engineering study.

The study analysis was conducted under the assumption that the  net refrigeration loads do not
vary with the outdoor ambient conditions, since they perform in  an air-conditioned indoor
environment. In reality, the refrigeration load in the display cases,  coolers, and freezers can vary
significantly during the year as a result of changes in the indoor  dry-bulb temperature and the
relative humidity. Capturing these variations and implementing them into the energy analysis
requires adequate performance data (mainly refrigerating load and evaporating temperature)
from the manufacturers of refrigerated fixtures. These data are generally not available and
require a large number of additional tests from the original equipment manufacturer. Obtaining
each data point by the currently used test method (ANSI/ASHRAE Standard 72 Method of
Testing Commercial Refrigerators and Freezers) is time-consuming. With the variety of
refrigerated fixtures and the pace of developing new models and improving the existing ones, it

[ 22 ]                                                         Final Report, September 2, 2008


-------
is unrealistic to expect data on refrigerating loads and evaporating temperatures as a function of
the dry and wet bulb indoor temperatures to become available in the near future. Yet, such data
would contribute substantially to improving the design and operation of efficient supermarket
refrigeration systems and to finding optimum design conditions minimizing the energy
consumption of the refrigeration and air-conditioning systems.

The load from the mechanical subcooling varies with outdoor ambient conditions.  This variation
is expressed through the enthalpy  of the liquid refrigerant entering the mechanical  subcooler and
was analyzed separately for each bin temperature and number of hours in the bin. Thus, the
cooling load from the mechanical  subcooler was determined from the following equation:
              QMSC = UlLTR (hcd,out - llMSCout),              (4)

where:

QMSC = Cooling capacity of the mechanical subcooler, Btu/hr

niLTR = LT refrigerant mass flow rate, Ib/hr

hcd!0ut= Specific enthalpy of the refrigerant at the condenser outlet, Btu/lb

hMscout= Specific enthalpy of the refrigerant at the mechanical sub-cooler outlet, Btu/lb

The LT refrigerant mass flow rate was calculated from the LT net cooling capacity and the
specific refrigeration capacity of the refrigerant at the outlet and inlet of the refrigerated
fixtures/evaporators applying the following equation:

                      QLT

               niLTR =  --------------------------- ,               (5)

                   llLTEvapOut - hLTMSCout

 where:

mLTR = LT refrigerant mass flow rate, Ib/hr

QLT = Cooling load in the LT system, Btu/hr

hLTEvaPout= Specific enthalpy of the  refrigerant exiting the refrigerated fixture, Btu/lb.

         = Specific enthalpy of the  LT refrigerant leaving the mechanical sub-cooler, Btu/lb.
An assumption was made that there will be no heat gains or losses in the liquid refrigerant lines
between the mechanical sub-cooler and refrigerated fixtures. (A detailed engineering study could
account for these heat gains or losses.)

The specific enthalpy of the refrigerant exiting a refrigerated fixture is determined at the
evaporating pressure and the temperature of the superheat vapor at the outlet of the LT fixture,
specified for each technology in Table 2.

The specific enthalpy of the refrigerant leaving the mechanical sub-cooler is determined at the
liquid refrigerant sub-cooled temperature for each technology (see Table 2).
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                      [ 23 ]


-------
The heat gains in the secondary-coolant supply and return lines and the heat gain from the
circulation pumps are additional loads added to the loads from the refrigerated fixtures. In this
study, the heat gains in the secondary-coolant supply and return lines have not been accounted
for but could be the subject of a future detailed engineering study. The heat load from the
circulation pumps is taken into consideration by adding to the particular secondary-coolant
circuit load, MT or LT, an estimated 90% of the power input into the electric motor of the
circulation pumps.

5.4. Bin energy consumption

The energy consumption (in kWh) in each bin was determined by multiplying the system power
per bin (in kW) times the number of hours in that bin.

5.5. Annual energy consumption

The total annual energy consumption is the total of the energy  consumption in all bins.
[24;
Final Report, September 2, 2008


-------
6.  RESULTS: BIN AND ANNUAL ENERGY CONSUMPTION OF
THE BASELINE AND ALTERATIVE TECHNOLOGIES

The results from the theoretical analysis are summarized in Table 8 and as a bar graph in Figure
5. Table 8 shows the annual energy consumption of the baseline and alternatives organized by
geographic location, technology, and system temperature level. Figure 5 shows the patterns in
the  annual energy consumption by refrigeration technology and geographic location. Appendix C
provides a more detailed set of results tables. Presented for the baseline and each alternative
within each geographical location, these tables illustrate how annual energy consumption was
calculated based on the power input and weather bin data.

As shown in Table 8, the results indicate that both secondary-coolant and distributed systems are
viable alternatives to the current centralized DX systems. All systems analyzed in the three
regions are within a few percent of the baseline in terms of energy use. Many other factors
regarding the actual operation of the systems are likely to lead to at least this amount of
fluctuation in energy use.

Boulder. In areas with a large number of hours with low ambient temperatures,  secondary-
coolant systems have the lowest annual energy consumption when liquid refrigerant is subcooled
to 30°F. In Boulder, the annual energy consumption for Alternative B (SC) was  4.1% lower than
the  DX baseline for systems with liquid refrigerants subcooled to 30°F, 3.2% lower for systems
subcooled to 40°F, and 2.4% lower for systems subcooled to 50°F. Distributed systems show
similar results as the secondary-coolant systems, with energy consumption for Alternative C
(DS) 3.3% lower than baseline energy consumption. As  shown in the table, for Alternative A
(MTS), which has a low-temperature DX and medium-temperature secondary-coolant
refrigeration system, annual energy consumption is 0.9% lower than Baseline (DX) system
energy use.

Philadelphia. In Philadelphia,  annual energy consumption was lowest for Alternative C (DS), at
3.3%  less than Baseline (DX) energy consumption. The Alternative B secondary-coolant systems
also resulted in reduced energy consumption, ranging from 0.8% to 2.5% lower  than the baseline
system. In Philadelphia, annual energy consumption for Alternative A (MTS) is 0.2% higher
than for the baseline system.

Atlanta. In areas with fewer hours of low temperatures,  distributed systems show the lowest
annual energy consumption. In Atlanta, Alternative C (DS) consumed 3.4% less energy than the
DX baseline system, while the  secondary-coolant systems consumed between 1.5% and 3.2%
more than the baseline. Annual energy consumption for Alternative A (MTS) is  3.1% higher than
for the baseline system.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                     [ 25 ]


-------
 Table 8: Annual energy consumption of supermarket refrigeration technologies at three
 geographic locations
                 System Type                       LT             MT            Combined Total
 Atlanta, GA Results
 Baseline DXwith 50°F(10°C) LT Liquid
 Alternative A MTSC/LTDX with 50°F(10°C)
 LT Liquid
 Alternative B SC with 50°F(10°C) LT Liquid
 Alternative B SC with 40°F(4°C)  LT Liquid
 Alternative B SC with 30°F(-1°C) LT Liquid
 Alternative C DS with 50°F(10°C) LT Liquid
   LT
 System
 Energy
kWh/Year


   339,627
   339,627

   339,838
   323,473
   307,964
   339,627
   MT
 System
 Energy
kWh/Year
     594,186
     623,416

     624,258
     632,425
     639,967
     562,871
 System
 Energy
kWh/Year

   933,813
   963,043

   964,096
   955,899
   947,931
   902,499
                                                                                           Compared
                                                                                             toDX
 3.1%

 3.2%
 2.4%
 1.5%
-3.4%
Boulder, CO Results
Baseline DXwith 50°F(10°C) LT Liquid
Alternative A MTSC/LTDX with 50°F(10°C)
LT Liquid
Alternative B SC with 50°F(10°C) LT Liquid
Alternative B SC with 40°F(4°C) LT Liquid
Alternative B SC with 30°F(-1°C) LT Liquid
Alternative C DS with 50°F(10°C) LT Liquid
Philadelphia, PA Results
Baseline DXwith 50°F(10°C) LT Liquid
Alternative A MTSC/LTDX with 50°F(10°C)
LT Liquid
Alternative B SC with 50°F(10°C) LT Liquid
Alternative B SC with 40°F(4°C) LT Liquid
Alternative B SC with 30°F(-1°C) LT Liquid
Alternative C DS with 50°F(10°C) LT Liquid

330,651
330,651

316,919
303,532
289,049
330,651

333,877
333,877

324,243
309,817
294,959
333,877

544,427
536,371

536,903
543,368
550,253
515,452

561,044
562,628

563,253
570,318
577,396
531,317

875,078
867,022

853,822
846,900
839,302
846,102

894,921
896,505

887,496
880,135
872,355
865,194

-
-0.9%

-2.4%
-3.2%
-4.1%
-3.3%

-
0.2%

-0.8%
-1.7%
-2.5%
-3.3%
 LEGEND:
 Baseline DXwith 50°F (10°C) Liquid: Baseline direct-expansion (DX) refrigeration system with min condensing
 temperature 70°F in both medium-temperature (MT) and low-temperature (LT) circuits, and LT liquid refrigerant
 subcooled to 50°F by mechanical subcooling in MT circuit.
 Refrigeration load from refrigerated fixtures: MT with saturation suction temperature (SST) 20°F 856,079
 BTU/h, LTwith SST-20°F 300,000 BTU/h. Alternative A MTSC/LTDX with 50°F (10°C) LT liquid: Refrigeration
 system with MT secondary-coolant (SC) circuit with 17°F SST and 22°F secondary-coolant supply temperature
 (SCST) and DX LT circuit with SST -20°F. Min condensing temperature 70°F in LT DX and 50°F in MT SC circuits.
 LT liquid refrigerant subcooled to 50°F by mechanical subcooling in MT circuit. Refrigeration load from refrigerated
 fixtures: MT 856,079 BTU/h, LT 300,000 BTU/h.
 Alternative B SC with 50°F (10°C) LT Liquid: SC refrigeration system with min condensing temperature 50°F in
 MT and 40°F in LT circuits. LT liquid refrigerant subcooled to 50°F (when condensing temperature is above 50°F)
 by mechanical subcooling in MT circuit. SCST 22°F in MT and -18°F in LT circuits. Refrigeration load from
 refrigerated fixtures: MT 856,079 BTU/h, LT 300,000 BTU/h.
 Alternative B SC with 40°F (10°C) LT Liquid: SC refrigeration system with min condensing temperature 50°F in
 MT and 40°F in LT circuits. LT liquid refrigerant subcooled to 40°F (when condensing temperature is above 40°F)
 by mechanical subcooling in MT circuit. SCST 22°F in MT and -18°F in LT circuits. Refrigeration load from
 refrigerated fixtures: MT 856,079 BTU/h, LT 300,000 BTU/h.
 Alternative B SC with 30°F (10°C) LT Liquid: SC refrigeration system with min condensing temperature 50°F in
 MT and 40°F in LT circuits. LT liquid refrigerant subcooled to 30°F by mechanical subcooling in MT  circuit. SCST
 22°F in MT and -18°F  in LT circuits.
 Refrigeration load from refrigerated fixtures: MT 856,079 BTU/h, LT 300,000 BTU/h. Alternative C DS with
 50°F (10°C): Distributed DX refrigeration systems with min condensing temperature 70°F in both MT and LT.
 Refrigeration load from refrigerated fixtures: MT with SST 25°F 450,000 BTU/h, MT with SST 20°F 406,079 BTU/h,
 LT with SST -20°F 300,000 BTU/h. LT liquid refrigerant  subcooled to 50°F by  mechanical subcooling in the
 adjacent MT circuit with SST 25°F.	
[26;
                       Final Report, September 2, 2008


-------
Figure 5: Annual energy consumption of supermarket refrigeration technologies in three
geographic locations
Total System Energy Comparison of DX Baseline vs. Alternative
Technologies
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7. ANALYSIS OF THE RESULTS
The focus of this theoretical study primarily involved an energy analysis of alternative
supermarket refrigeration technologies as compared to a baseline DX technology. Some previous
studies of the energy efficiency of alternative supermarket refrigeration systems, particularly
secondary-coolant technologies, have shown secondary-coolant refrigeration systems to be
associated with up to 30% higher annual energy consumption compared to DX systems.
However, these studies have involved a limited number of secondary coolants with poor thermo-
physical properties, a lack of a good design practice, and in some instances, design errors. For
this reason, this study represents an attempt to conduct an analysis based on the most advanced
design practices and using secondary coolants with improved performance properties.

Annual energy consumption is a reliable indicator of the design and operational efficiency of a
supermarket refrigeration system. When comparing the three alternative technologies with the
baseline, it becomes apparent that no one technology will be superior in all geographic locations
in terms of energy efficiency.

In climates with fewer hours of low annual ambient temperatures, such as Atlanta, GA,
Alternative C (DS) distributed systems have the lowest annual energy consumption by 3.4%. In
comparison, Alternative B (SC) systems have between 1.5% and 3.2% higher energy
consumption than the baseline. The design features of the distributed systems lead to the
conclusion that the two temperature levels in the MT load (+20°F and +25°F) have contributed
to the high efficiency of Alternative C. Because of the prevailing size of the MT load, which is
approximately three times the size of the LT load, any efficiency-improving measure in MT will
have a noticeable impact on the annual energy consumption of the whole system. The same
efficiency-enhancing effect can be  achieved in the other technologies by using multiple suction
groups, which are analogous to the multiple temperature levels in Alternative C.

A second conclusion is that multiple suction groups in any technology have the most significant
impact in geographic locations with warmer climates. In such climates, the special features of the
secondary-coolant technologies, such as the lower limit of the floating condensing temperature
and the deeper mechanical subcooling of the LT liquid refrigerant cannot make up for the
benefits from the multiple MT suction groups because in milder climates these special features
cannot materialize their full potential. In warmer climates, the use of a complete secondary-
coolant technology (Alternative B) can be counterproductive from an energy point of view. This
situation can be exacerbated when a secondary-coolant technology is applied only to the MT
system (e.g., Alternative  A), preventing the implementation of multiple suction groups or
distributed systems.

The benefits from the special features of the alternative technologies have a different relative
impact in geographic locations with a larger number of hours of low ambient temperatures. In
Boulder, CO, the version of the secondary-coolant technology (Alternative B - SC) with a level
of liquid refrigerant subcooling of 30°F has the lowest energy consumption. Since subcooling to
30°F has become the norm in the design practice of at least one major original equipment
manufacturer, the secondary-coolant technology can be expected to have low energy

[ 28 ]                                                        Final Report,  September 2, 2008


-------
consumption in geographic locations with climates similar to Boulder, Co.. Apparently, the
lower annual energy consumption in the secondary-coolant technology in climates with a larger
number of low ambient temperatures results from the lower limit of floating condensing
pressure/temperature and lower subcooling. Because of the larger number of hours with low
ambient temperatures, both LT and MT compressors operate longer at low discharge pressures
and consume less energy.  Some supermarket industry experts have suggested that the baseline
DX systems can be operated at the same low limit of the condensing pressures as Alternative B
(SC), and energy savings could be expected if DX systems were operated in this manner.
However, secondary-coolant systems are especially suitable for low condensing pressures and
low liquid subcooling because of the short liquid refrigerant lines upstream from the expansion
valves.

The energy benefits of Alternative C (DS) in low-ambient climates are similar to the benefits
from Alternative B (SC), due to the  multiple temperature levels or multiple suction groups in the
DS system. Thus, the decision of which system to select may depend on consideration of other
issues, such as ease and cost of operation and maintenance, the supermarket's established
practices and preferences, and installed cost.

In the climate conditions of Philadelphia, PA, the only alternative technology that did not use
less energy than the baseline system was Alternative A. The comparable annual energy
consumption between Alternative B with 30°F subcooled liquid and Alternative C indicates that
the decision of which technology to choose will depend on  additional considerations.

The interpretation of the results becomes even more evident from the number of hours the MT
and LT compressors operate at their minimum SDT at the three geographic locations (see Table
9). The MT compressors will operate at their minimum SDT (50°F) 2.3 times longer in Boulder,
CO and 2.2 times longer in Philadelphia, PA as compared to Atlanta, GA. The LT compressors
will operate at their minimum SDT (40°F) 4.0 times longer in Boulder, CO and 2.8 times longer
in Philadelphia, PA as compared to Atlanta, GA. Therefore, technologies that can operate the
compressors at the lowest SDT are expected to have  a prevailing energy efficiency benefit in
geographic areas with climates similar to or colder than Boulder, CO and Philadelphia, PA. Their
energy efficiency advantage is expected to be negligible or non-existent in geographic locations
with climates similar to or warmer than Atlanta, GA.

To summarize, the results of the analysis of alternative supermarket refrigeration technologies at
the three  geographic locations indicate that two of the three analyzed alternative technologies
have lower energy requirements than the baseline DX technology in these climates.  Multi-
temperature distributed systems (Alternative C)  are the best choice in climate conditions such as
Atlanta, GA or warmer. Secondary-coolant technologies (Alternative B) and distributed systems
(Alternative C) provide energy benefits in climate conditions such as Philadelphia, Boulder, or
colder. The third technology, Alternative A, only showed energy advantages compared to the
baseline DX system in Boulder, CO. In all three locations, Alternative A showed energy
penalties of up to a few percent compared to the other alternative technologies.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                     [ 29 ]


-------
Table 9: Number of hours of MT and LT compressors at their minimum operating SDT (50°F
for WIT and 40°F for LT) at the three geographic locations
Ambient
Temp.
Bin
°F


35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Subtotal
Relative
to Atlanta
MT
Compr.
Min.
Cond.
Temp.
°F
50
50
50
50
50
50
50
50
(hours):

(ratio):
a The number of hours
integrating the 5°F bin
performance
Weather
Bin Data
Atlanta,
GA
Hours

560
323
181
72
64
7
0
0
1207

1
in this bin is
and the next
Weather
Bin Data
Boulder,
CO
Hours

834
717
611
251
201
130
89
126a
2833

2.3
the cumulative
Weather
Bin Data
Philadel-
phia, PA
Hours

1067
685
442
248
184
40
0
0
2666

2.2
Ambient
Temp.
Bin
°F


35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5



LT
Compr.
Min.
Cond.
Temp.
°F



40
40
40
40
40
40



Weather
Bin Data
Atlanta,
GA
Hours



181
72
64
7
0
0
324

1
Weather
Bin Data
Boulder,
CO
Hours



611
251
201
130
89
126a
1282

4.0
Weather
Bin Data
Philadel-
phia, PA
Hours



442
248
184
40
0
0
914

2.8
number of hours of all temperatures within and below the 5°F-bin, thus
lower-temperature bins. The
of the refrigeration system in a similar way and can
reason is that
be processed
all these temperatures
together.

affect the



[so;
Final Report, September 2, 2008


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8.  SUMMARY OF CONCLUSIONS AND RECOMMENDATIONS
FOR NEXT STEPS

8.1. Summary of conclusions

A general conclusion from this analysis is there are viable alternative supermarket technologies
with equal or better energy efficiency to the baseline DX technology. Depending on geographic
location, the alternative technology of choice is either a secondary-coolant (Alternative B) or a
distributed (Alternative C) system.

In geographic areas with a large number of hours with ambient temperatures below 40°F, the MT
compressors will operate at their lowest allowable SDT (50°F) with reduced energy
consumption. At ambient temperatures below 30°F, the LT compressors will also operate at their
minimum allowable SDT (40°F) with reduced energy consumption. The prolonged operation of
both MT and LT systems with low energy consumptions  in geographic areas with such ambient
conditions will lead to a lower annual energy consumption of the SC refrigeration systems
compared to the baseline. The distributed systems show a similar level of energy performance in
these cold climates.

In geographic areas with a limited number of hours below 40°F, the secondary coolant systems
do not have competitive annual energy consumption. The most advantageous technology for
these conditions is Alternative C (distributed refrigeration systems), with as many SST levels as
feasible with respect to installed  cost.

In geographic areas with ambient conditions falling between the two climate extremes studied
here, both alternative technologies, Alternative B (secondary-coolant) and Alternative C
(distributed), offer about equal energy efficiency  and the  choice between these technologies will
reflect additional considerations (such as ease and cost of operation and maintenance, the
supermarket's established practices and preferences and installed cost).

The conclusions in this study are supported by the practices in some of the major supermarket
chains operating in the northeastern and southeastern states. Secondary-coolant systems have
become the exclusive technology for a large supermarket chain in the northeast. In addition to
the measurable lower annual energy use compared to other alternatives, lower operating costs
have been reported, due to low or no maintenance, low or no loss of refrigerant, lower shrinkage,
and better product quality.2

Another large supermarket chain operating in the southeast achieves favorable annual  energy
consumption by using multiple suction groups in  its DX systems.3 In this case, distributed
systems would reduce the amount of refrigerant charge while maintaining the same energy
efficiency. In addition, a large national chain has  initiated aggressive cost- and energy-cutting
2 FMI Energy and Technical Services Conference, Miami, FL, September 2002.
3 Confidential materials submitted by a supermarket chain.

Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                     [ 31 ]


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measures through the deployment of optimized distributed systems. Whenever justified, this
chain also deploys secondary systems.4

An important conclusion of this study is that no one technology has competitive annual energy
consumption in all climate conditions. When planning a new store in a different location, it is
important to estimate the annual energy consumption for all technologies under consideration.
Some of the other factors to consider are:

   •   Cost of equipment
   •   Cost and ease of installation
   •   Refrigerant and secondary coolant costs
   •   Cost and ease of operation and maintenance
   •   Other performance issues (e.g., food quality and shrink).

8.2. Recommendations for next steps

A large number of factors affect the performance and annual energy consumption of a
refrigeration system. This theoretical study was performed based on a number of simplifying
assumptions in order to provide a preliminary assessment of the feasibility of alternative
supermarket refrigeration technologies based on conditions that reflect some of the recent
advancements in the alternative and baseline technologies, and to determine if a more detailed
engineering study, involving a higher level of effort, is needed to more fully analyze the
alternative systems.  The results from this study indicate that two  of the investigated alternative
technologies, Alternative B (secondary-coolant) and Alternative C (distributed), are viable DX
alternatives and that a more detailed engineering study could provide data that are more accurate
and more closely reflect the real systems and practices, including recent advancements, for both
the baseline and the alternative technologies.

An expanded engineering study could include some or all of the following approaches:

   •   Conduct an engineering-based study that incorporates additional parameters and
       conditions that more accurately define currently available DX, SC, and DS supermarket
       refrigeration systems. This theoretical study was based on several simplified
       assumptions: 1) a limited number of suction groups, temperature levels,  and secondary-
       coolant supply temperatures, 2) omission of power input into condensing fans, 3)
       omission of heat gains and losses into refrigerant supply and return lines, and 4) omission
       of heat gains into secondary-coolant supply and return lines. These factors should be
       included in a detailed engineering study. Table 10 presents a summary of the key
       parameters and conditions to include in a more detailed engineering study. Appendix B
       contains a more detailed list of these factors.

   •   Evaluate the energy impact of the lower limit of floating condensing temperatures in a
       DX system.

   •   Consider the seasonal variation in fixture refrigeration loads.
4 Based on confidential conversations with a supermarket chain.

[ 32 ]                                                         Final Report, September 2, 2008


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   •   Conduct an investigation of a hybrid distributed/secondary-coolant technology, which
       could prove to be a successful combination of the benefits of the distributed and
       secondary-coolant systems.

   •   Include a secondary-coolant technology with a phase-change secondary fluid, in
       particular CC>2.

   •   Assess CC>2 as a primary refrigerant in a low-temperature cascade system.

   •   The dependency of the annual energy consumption on climate conditions justifies the
       expansion of a study that investigates additional geographic locations. A larger number of
       analyzed locations can become a building block for a technology map that will provide
       preliminary information on the suitability of each technology. Supermarkets could use
       this information during the planning process for building a new supermarket or
       remodeling an existing one to assess the viability of different technologies.

Proposals for parameters to study in a detailed engineering analysis are provided in Appendix B.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
;33]


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Table 10: Conditions for a detailed engineering analysis
Technology
Baseline




Alternative A
MTSC, LTDX




Alternative B
MTSC, LTSC
with Dynalene


Alternative C

DISTRIBUTED





System
Type
DX
DX
DX
DX
DX
DX
DX
SC

SC
SC
SC
SC
SC
SC
SC
DS
DS
DS
DS
DS
DS
DS
DS
Temp.
Level
LT
LT
MT
MT
MT
LT
LT
MT

MT
MT
LT
LT
MT
MT
MT
LT
LT
LT
MT
MT
MT
MT
MT
Unit#
a
1
2
3
4
4
1
2
3

4
4
1
2
3
4
4
1
6b
2
3
5
4a
6a
4b
Notes
Subcool Load to Unit 3
Subcool Load to Unit 3



Subcool Load to Unit 3
Subcool Load to Unit 3




Subcool Load to Unit 3
Subcool Load to Unit 3
Same as Alternative A
Same as Alternative A
Same as Alternative A
Subcool Load to Unit 3
Subcool Load to Unit
6a
Subcool Load to Unit 3





SSTb
°F
-25
-14
24
20
15
-25
-14
21

17
12
-27
-16
21
17
12
-24
-20
-13
24
24
20
20
15
Max SDT c
°F
110
110
115
115
115
110
110
115

115
115
110
110
115
115
115
110
110
110
115
115
115
115
115
MinSDTd
°F
70
70
70
70
70
70
70
50

50
50
40
40
50
50
50
50
50
50
50
50
50
50
50
Liquid Temp.
°F
50
50
SDT-56
SDT -5
SDT-5
50
50
SDT-5

SDT-5
SDT-5
50 (SCT-5)e,30
50 (SCT-5) e,30
SDT-5
SDT-5
SDT-5
50,45'
50,45 f
50,45 f
SDT-5
SDT-5
SDT-5
SDT-5
SDT-5
Case/Chiller Outlet9
°F
-6
5
39
35
30
-6
5
26

22
17
-22
-11
26
22
17
-6
-2
5
39
39
35
35
30
RGT
°F
Cooling load
Btu/hr
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain
SLHE

SLHE
SLHE
SLHE
SLHE
SLHE
SLHE
SLHE
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain
Add Heat Gain

Add MSC of units! &2




AddMSCu's1&2&
PH
AddPH
AddPH
AddPH
Add PH
Add MSCu's1&2&
PH
AddPH
AddPH



Add MSC of units! &2


Add MSC of unit 6b

Power
kW







add PP

addPP
addPP
addPP
add PP
add PP
addPP
addPP








MSC = Mechanical subcooling, PH = Pump heat, PP = Pump power, SDT = Saturation discharge temperature, RGT = Return gas temperature

a See Appendix B (Tables 1-10 and Figures 2-3) fora more detailed illustration of how these systems are configured.
b Clarify/confirm with GreenChill Technical Review Committee members the pressure drops for oil return in LTDX and MTDX. While 2°R for both LT&MT DX were assumed in the theoretical analysis, in
the current table for a detailed engineering analysis 2°R in MTDX and 3°R in LTDX equivalent pressure drop has been assumed.
0 Clarify/confirm with GreenChill Technical Review Committee members condenser sizing. While the theoretical analysis was performed for temperature difference 10.0°R for both LT & MT
condensers, the current table fora detailed engineering analysis assumes 10°R for LT and 15°R for MT condensers.
d Clarify/confirm with GreenChill Technical Review Committee members the minimum SDT for Alternative  C. While the theoretical analysis was performed for minimum 70°F, the current table for a
detailed engineering analysis assumes 50°F for both MT and LT.
e Clarify/confirm with GreenChill Technical Review Committee members the natural subcooling in the condensers. While the theoretical analysis was performed with no subcooling, the current table for
a detailed engineering analysis assumes 5°R natural subcooling in both LT and MT condensers.
f 50°F out of mechanical subcooler or SCT - 5 = min cond. - 5°R natural SC
9 Clarify/confirm with GreenChill Technical Review Committee members the superheat out of MT and LT display cases and intermediate heat exchangers (evaporator/chillers). While the theoretical
analysis was performed at 15°R superheat out of LT display cases, 5°R out of MT display cases, and 10°R out of both LTand MT intermediate heat exchangers, the current table fora detailed
engineering analysis assumes 19°R superheat out of LT DX display cases (3°R in the coil and 16°R in the suction/liquid heat exchanger), 15°R out of the MT DX display cases; 18°R out of LT DS
display cases, 15°R out of the MT DS display cases;  and 5°R superheat in both LTand MT intermediate heat exchangers (evaporator/chillers).
[34;
   Final Report, September 2, 2008
         www.epa.gov/greench i11
Chillin' for the environment

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Appendix A


Theoretical Analysis of Alternative Supermarket
Refrigeration Technologies:
Technical Review Committee Members
www.epa.gov/greenchill
Chiltin' for the environment

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 Theoretical Analysis of Alternative Supermarket Refrigeration Technologies:

                     Technical Review Committee Members
GreenChill Partners

Bob Garrity
Senior Vice President, Store Planning,
   Construction & Conservation
Giant Eagle, Inc.

Harrison Horning
Energy Manager
Hannaford and Sweetbay

Chris LaPietra
Wholesale Marketing Manager
Honeywell

Kathy Loftus, CEM
National Energy Manager
Whole Foods Market
North Atlantic Regional Office

Scott Martin
Director Sustainable Technologies
Hill PHOENIX, Inc.

Wayne Rosa
Strategic Sourcing Manager for Energy &
   Maintenance
Food Lion, LLC

Stephen Sloan
Refrigeration / Energy Program Manager
Publix Super Markets, Inc.
EPA

Julius Banks
Team Leader, Refrigerant Recovery and
   Recycling
Stratospheric Protection Division

Cynthia Gage, PhD.
National Expert, Senior Research Engineer
Office of Research and Development

David S. Godwin, P.E.
Environmental Engineer
Stratospheric Protection Division

Bella Maranion
Sector Analyst
Stratospheric Protection Division
[A-2;
             Final Report, September 2, 2008


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Appendix B
 EPA Supermarket Alternatives Study Report
 (August 6, 2007)

 Phase 1: Proposal for a detailed engineering
 analysis—description of a baseline store and
 alternative configurations
www.epa.gov/greenchill
Chiltin' for the environment

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         EPA Supermarket Alternatives Study
 Phase 1: Proposal for a detailed engineering analysis—description
         of a baseline store and alternative configurations
                         Prepared by:
                      Dr. Georgi Kazachki
                        CRYOTHERM
                     1442 Wembley Ct. NE
                       Atlanta, GA 30329
                        August 06, 2007
                 (Introduction Revised December 19, 2007)
[B-2;
Final Report, September 2, 2008


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CONTENTS:

1.   Introduction
2.   Parameters affecting the performance and energy efficiency of a supermarket refrigeration system:
       2.1. Summary of the parameters for energy comparison
               2.1.1. Systems to be investigated
               2.1.2. Store size, location, and assumptions
               2.1.3. Conditions for the analysis
       2.2. Piping diagrams for the baseline and alternative configurations
3.   Definition of the baseline store:
       3.1. Floor plan and location of the refrigeration loads
       3.2. Load  distribution, load components, and piping
4.   Definition of Alternative A:
       4.1. Location of the refrigeration loads.
       4.4. Load  distribution, load components, and piping
  5.   Definition of Alternative B:
       5.1. Location of the refrigeration loads.
       5.4. Load  distribution, load components, and piping
  6.   Definition of Alternative C:
       6.1. Location of the refrigeration loads.
       6.4. Load  distribution, load components, and piping
    7. Ambient dry-bulb temperatures for Atlanta, Boulder, CO, and Philadelphia
    8. Illustrations:
    Table 1: DX Baseline  LT Unit  #1, -25 °F, Refrigerant HFC-404A
    Table 2: DX Baseline  LT Unit  #2, -14°F Refrigerant HFC-404A
    Table 3: DX Baseline  MT Unit #3, +24°F, Refrigerant HFC-404A
    Table 4: DX Baseline  MT Unit #4, +20°F/+15°F, Refrigerant HFC-404A
    Table 5: Distributed DS-1, -25°F, Refrigerant HFC-404A
    Table 6: Distributed DS-2, -14°F, Refrigerant HFC-404A
    Table 7: Distributed DS-3, +24°F, Refrigerant HFC-404A
    Table 8: Distributed DS-4a +20°F and DS-4b +15°F, Refrigerant HFC-404A
    Table 9: Distributed DS-5, +24°F, Refrigerant HFC-404A
    Table 10: Distributed DS-6a, +20°F and DS-6b, -20°F, Refrigerant HFC-404A
    Table 11: Ambient Dry-Bulb Temperatures for Atlanta, GA
    Table 12: Ambient Dry-Bulb Temperatures for Boulder, CO
    Table 13: Ambient Dry-Bulb Temperatures for Philadelphia, PA

    Figure 1: Piping schematics of the baseline and alternative systems
    Figure 2: Fixture plan DX baseline
    Figure 3: Fixture plan Distributed System
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                      [ B-3 ]


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1. INTRODUCTION

EPA is developing a voluntary partnership with the supermarket industry to facilitate the transition from
ozone-depleting substances to ozone-friendly alternatives. Known as the GreenChill Advanced
Refrigeration Partnership, the overall goal of this activity is to promote the adoption of technologies,
strategies, and practices that lower emissions of ozone-depleting substances (ODS) and greenhouse gases
(GHGs) through both the reduction of refrigerant emissions and the increase of refrigeration systems'
energy efficiency. One aspect of the partnership is to conduct technological research and share
information that will aid partners in meeting the GreenChill goals.

To meet this goal, EPA commissioned a study to compare the energy efficiency of alternative
supermarket refrigeration technologies. The study, Theoretical Analysis of Alternative Supermarket
Refrigeration Technologies, is based on a theoretical analysis of the energy efficiency of the three most
common technologies:

    •   Direct-expansion (DX) centralized systems,

    •   Secondary-loop, secondary-coolant, centralized systems, and

    •   Distributed systems.

The analysis is based primarily upon existing thermodynamic and heat transfer data for refrigerants and
secondary-coolant fluids, and performance characteristics  from existing laboratory and/or field
measurements, manufacturer data,  or other available information. The study assesses the following four
supermarket refrigeration scenarios:

Baseline:      New supermarket with a DX refrigeration system using an HFC refrigerant (DX).
Alternative A: New supermarket with a Low Temp DX and Medium Temp glycol  secondary loop
               refrigeration system using an HFC refrigerant (MTS).
Alternative B: New supermarket with a secondary loop refrigeration system using  an HFC refrigerant
               (SC).
Alternative C: New supermarket with a distributed refrigeration system using an HFC refrigerant (DS).

This Phase 1 report represents the first phase of the theoretical study. It involved a series of conference
calls with the GreenChill Technical Review Committee and EPA to scope out the parameters and
methodologies that could be used to estimate annual energy use of various types of supermarket
refrigeration systems. The resulting Phase 1 report describes parameters and methodologies that were
developed from this process. Upon consideration, it was determined that these parameters were
appropriate for conducting a detailed engineering analysis of the annual energy use of the baseline and
alternative systems, rather than a simplified theoretical study that reflects currently-designed supermarket
refrigeration systems. Consequently, the proposed parameters and assumptions were simplified for the
theoretical study (for example, the  theoretical study is based on fewer suction groups than suggested in
this Phase 1 report - see Chapter 4  of the main report).

This Phase 1 report describes the proposed engineering study that was initially developed. This could
provide the basis for follow-on work to the existing theoretical study.
[B-4;
Final Report, September 2, 2008


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2. PARAMETERS AFFECTING THE PERFORMANCE AND ENERGY EFFICIENCY
OF A SUPERMARKET REFRIGERATION SYSTEM

The major parameters are:
    •   Store location
    •   Indoor data
    •   Refrigeration loads
    •   Suction saturation temperature
    •   Discharge saturation temperature
    •   Liquid refrigerant subcooling
    •   Refrigerant vapor superheat
    •   System design:
           o   Type of system
           o   Refrigerant selection
           o   Secondary coolant selection
           o   Components selection
           o   Tailoring the system to the refrigerant properties

2.1. Summary of the  parameters for energy comparison:

2.1.1.  Systems to be investigated:

Baseline:      New supermarket with a DX refrigeration system using an HFC refrigerant (DX).
Alternative A:  New supermarket with a Low Temp DX and Medium Temp glycol secondary loop
               refrigeration system using an HFC refrigerant (MTS).
Alternative B:  New supermarket with a secondary loop refrigeration system using an HFC refrigerant
               (SC).
Alternative C:  New supermarket with a distributed refrigeration system using an HFC refrigerant (DS).

2.1.2.  Store size, location, and assumptions:

1.   Baseline store will be 45,000 sq. ft. with HFC-404A.
2.   Locations will be Atlanta, Philadelphia and Boulder, CO.
3.   Heat reclaim and defrost method will be excluded from the analysis.
4.   Heating and air-conditioning loads, building fire and safety code, store lighting, plug loads and other
    loads, HVAC annual consumption will be excluded from this study.
5.  Note: To avoid the effects of compressor designs, models, cycling, and control strategies, the analysis
    for the base line and all alternatives will use the energy efficiency ratio (EER) of a representative
    compressor based on manufacturer's data calculated at the required operating conditions of each
    alternative technology rather than selecting individual compressors for each alternative technology.
6.  Note: to avoid the effect of the compressor design on the technology comparison, the use of scroll
    compressors with EVI needs to be a subject of another study. Since scroll compressors with EVI can
    be used in the baseline and in all alternatives, their potential use will equally impact all technologies.

2.1.3.  Conditions for  the analysis:

The analysis will be performed at the following conditions:
1.  Number of the distributed groups for Alternative C:
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
:B-S]


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           a.   Three saturation suction temperatures for LT (-25, -20, -15°F) and three saturation
               suction temperatures for MT (+24, +20, and +15°F) located strategically on the roof
               above the associated line-ups.
           b.   The 6 suction saturation temperatures will be distributed among 8 groups in 6 locations.
2.  Use of suction-line-liquid-line heat exchanger (SLHX) in the display cases. Since both, presence and
    absence of SLHX, are observed, and the SLHX size and efficiency vary by case manufacturers, the
    analysis will be performed with superheat out of the display cases equal to the superheat at the coil
    exit plus additional 5R for MT and additional 10R for LT regardless whether this has resulted from
    SLHX or through direct heat transfer between the air inside the display case and the suction line
    between the coil outlets and the case outlet.
3.  Use of SLHX on the rack in Alternative B  (SC) - optional.
4.  Exit superheat from the evaporators and display cases:
        a.  Exit superheat for MT evaporators: 8R
        b.  Exit superheat for LT evaporators: 6R
        c.  Exit superheat for MT display cases: 13R
        d.  Exit superheat for LT display cases: 16R
    The superheat increase of 5R in the MT and 10R in the LT display cases are to account for possible
    use of SLHE or other similar useful superheat between the evaporator and the case outlet.
5.  Mechanical subcooling (MS)  of the LT liquid refrigerant by the MT refrigerant:
        a.  In Baseline, to 50°F
        b.  In Alternative  A (MTS), to 50°F.
        c.  In Alternative B (SC), to 50°F and 30°F.
        d.  In Alternative  C (DS), to 50°F.
6.  Impact of heat gains/losses in the liquid refrigerant lines on subcooling at the display cases and
    intermediate heat exchanger (IHX):
        a.  In Baseline and LT line of Alternative A, the liquid temperature will increase as a result of
        the heat gains. The increase will be calculated from the diameters,  lengths, and insulation of the
        liquid lines.
        b.  In Alternative  C (DS), the heat losses will be calculated from the diameters, lengths, and
        insulation of the liquid lines.
        c.  In Alternative B (SC) and MT line of Alternative A, the increase of the liquid refrigerant
        temperature can be neglected because of the short liquid lines.
7.  Heat gains in SC supply and return lines in Alternative B (SC) and MT line of Alternative A (MTS)
    will be calculated from the SC properties, temperatures, and geometry  (diameters, lengths, and
    insulation) in the MT and LT circuits. The  heat gains will be added to the cooling load of the display
    cases.
8.  Temperature difference (TD)  between ambient-air temperature and condensing temperature will be
    used rather than type of condensers (air-cooled, evaporative, or water-cooled), manufacturers and
    model numbers. Condenser TD:
        a.  Medium-temperature system 15R
        b.  Low-temperature system 10R
9.  Natural subcooling in the condensers: 5R for all systems.
10. Condenser fan control:
        a.  In Baseline, float SDT to 70°F for MT and LT condensers.
        b.  In Alternative A (MTS), float SDT to 50°F for MT and to 70°F for LT condensers.
        c.  In Alternative B (SC), float SDT to 50°F for MT and to 40°F for LT condensers.
        d.  In Alternative C (DS), float SDT to 50°F  for both MT and LT condensers.
11. Condenser fan consumption:  consider it by fan kW/THR for all technologies.
        Note: THR = Total Heat Rejection, BTU/hr
12. MT saturation suction temperature (SST) in Alternative A (MTS) and  both MT and LT SST in
Alternative B (SC) to be 3R lower than the corresponding SST in the DX suction groups in Baseline.

[  B-6 ]                                                             Final Report, September 2, 2008


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       Note: This results from the assumed 5R temperature difference in the MT and LT intermediate
       heat exchangers (IHX) and the absence of2R equivalent pressure drop in the suction lines for oil
       return.
14. Compressor inlet pressure:
       a.  Pressure drop in DX MT suction lines: 2R equivalent
       b.  Pressure drop in DX LT suction lines: 2R equivalent
       c.  Pressure drop in DS MT suction lines: 2R equivalent
       d.  Pressure drop in DS LT suction lines: 2R equivalent
       e.  Pressure drop in Alternatives A (MTS) and B (SC) suction lines: equivalent of less than 0.5R
       lower than the IHX evaporating temperature because of the short return lines and the downstream
       movement of oil.
15. Compressor inlet temperature:
       a.  In Baseline, Alternatives  A (LT line), and C (DS), the compressor inlet temperature will be
       equal to the temperature at the outlet of the display cases plus temperature increase from the heat
       gains in the return lines. These will be calculated.
       b.  In Alternative B, the temperature increase from heat gains will be neglected because of the
       short lines.
16. Secondary-coolant supply/return temperature difference: 6, 8, and 10R
17. Circulation pumps:
       a.  The power input into the  SC circulation pumps will be added to the power input of the
       compressor racks.
       b.  The heat from the pumps will be added to the cooling load of the racks.
18. Compressors: in the report the compressor manufacturer and compressor models will be blanked out.
The same applies for any information that may be perceived as biased.
19. Refrigerant R-404A will be used  in the study.
20. Analysis with both Dynalene and CO2 as a secondary coolant in Alternative B LT loop.
21. In Alternative A (MTS) and Alternative B (SC), glycol will be used in the MT loop.
22. Indoor temperature and relative humidity for the study: 75/55% year around.
23. Insulation - Rubatex with thickness:
       a.  MT DX: liquid !/2", suction 3/4"
       b.  LT DX: liquid 3/4", suction 1"
       c.  MT SC supply and return: 1"
       d.  LT SC supply and return: l!/2"
       e.  MT DS: liquid l/2\ suction %"
       f  LT DS liquid 3/4", suction 1

2.2. Piping diagrams for the baseline and alternative configurations

Schematics of the baseline and alternative configurations are presented in Figure 1.

3. DEFINITION OF THE BASELINE STORE:

       3.1. Floor plan and location of the refrigeration loads - Figure 2.
       3.2. Load distribution,  load components, and piping - Table 1 to 4.

4. DEFINITION OF ALTERNTATIVE A

       4.1. Location of the refrigeration loads - same as for the baseline.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies                      [ B-7 ]


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       4.2. Load distribution, load components, and piping - Load distribution and components are the
       same as for the baseline. The piping for the LT system is the same as for the baseline. The piping
       for the MT system will be determined in the second phase of the project.

5. DEFINITION OF ALTERNATIVE B
       5.1. Location of the refrigeration loads - same as for the baseline.
       5.2. Load distribution, load components, and piping - Load distribution and components are the
       same as for the baseline. The piping for the LT and MT systems will be determined in the  second
       phase of the project.
6. DEFINITION OF ALTERNATIVE C:
       6.1. Location of the loads and units - Figure 3.
       6.2. Load distribution and load components - Table 5 to 10. The piping for the LT and MT
       distributed systems will be determined in the second phase of the project.

7. AMBIENT DRY-BULB TEMPERATURES
Ambient dry-bulb temperatures that will be used in the analysis for Atlanta, Boulder, CO, and
Philadelphia are presented in Tables 11 to 13.
[B-s;
Final Report, September 2, 2008


-------
Table 1: DX Baseline LT Unit #1, -25°F, Refrigerant HFC-404A
DX Header Loads
                                    Load Line Sizes
# Loads

1
2
3
4
5
6
7
8
9
10
11
12
ID

SP
54
52
6
5
4
21
20
19
18
29
30
Load Description Model
Unit #1 Circuit Manifold Remote
Spare

10'x12'x10' Bakery/Deli Frzr, R=7/8
12'+(1)E Fz Island Case, R=7/8
12' Frozen Island Case, R=7/8
12'+(1)E Fz Island Case, R=7/8
10 Drs Ice Cream Cases, R=1
10 Drs Ice Cream Cases, R=1
10 Drs Ice Cream Cases, R=1
5 Drs Ice Cream Cases, R=7/8
16'x24'x10' 1C Freezer, R= 1-3/8

MBTU
129.6

8.0
9.3
10.5
7.6
10.5
14.1
14.1
14.1
7.0
20.8
136
Evap,°F
-22

-18
-18
-12
-12
-12
-20
-20
-20
-20
-22
-18
Run
40

377
322
170
190
202
210
230
270
290
172
79
Supply
7/8"

5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
Retur
2-5/8"

1-1/8"
1-1/8"
1-1/8"
7/8"
1-1/8"
1-3/8"
1-3/8"
1-3/8"
1-1/8"
1-5/8"
1-1/8"
Total Load #1 -25°F
MBTU
129.6
       DX Compressor Rack #1, Design conditions -25/110°F, Subcooled liquid temp, is SOT.
         Pos.  Compr. Model                  Capacity, MBTU  % Cap.     Rej.MBTU Rej. %
         1
         2
         3
         4
26.5
44.5
54.2
85.3
13%
21%
26%
41%
31.9
53.6
64.9
102.3
120%
120%
120%
120%
               Total Compressors Capacity
               Rack Capacity to Load Ratio
210.5
               162%
           252.7
120%
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-------
Table 2: DX Baseline LT Unit #2, -14°F, Refrigerant HFC-404A
DX Header Loads
# Loads ID
1 SP
2 40
3
4
5
6
7
8
9
10
11
12
13
14
Total






59
10
11
9
8
12
13
7
17
16
15
14
Load Description Model
Unit #2 Circuit Manifold Remote
Spare
3 Drs Frozen Fd Cases, R=5/8
3 Drs Frozen Fd Cases, R=5/8
10 Drs Frozen Fd Cases, R=1-1/8"
10 Drs Frozen Fd Cases, R=1-1/8"
10 Drs Frozen Fd Cases, R=1-1/8"
10 Drs Frozen Fd Cases, R=1-1/8"
10 Drs Frozen Fd Cases, R=1-1/8"
10 Drs Frozen Fd Cases, R=1-1/8"
10 Drs Frozen Fd Cases, R=1-1/8"
5 Drs Frozen Fd Cases, R=7/8"
10 Drs Frozen Fd Cases, R=1-1/8"
10 Drs Frozen Fd Cases, R=1-1/8"
10 Drs Frozen Fd Cases, R=1-1/8"
Load#2-14°F MBTU
DX
Pos.
1
2
3
4
Compressor Rack #2, Design conditions -14/110°F,
Compr. Model Capacity, MBTU
42.0
50.7
60.0
82.5
MBTU
149.7
4.0

13.5
13.5
13.5
13.5
13.5
13.5
13.5
6.7
13.5
13.5
13.5
149.7
Subcooled
% Cap.
18%
22%
26%
35%
Evap,°F
-11
-11

-11
-11
-11
-11
-11
-11
-11
-11
-11
-11
-11

liquid temp.
Rej.MBTU
47.5
57.4
68.1
93.5
Run
40
120
322
230
220
220
170
170
150
150
245
235
185
165
is 50°F.
Rej. %
113%
113%
114%
113%
Sup
7/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"






                                     Load Line Sizes
                                                      Return
                                                      2-5/8"

                                                      7/8"
                                                      7/8"
                                                      1-3/8"
                                                      1-3/8"
                                                      1-3/8"
                                                      1-3/8"
                                                      1-3/8"
                                                      1-3/8"
                                                      1-3/8"
                                                      7/8"
                                                      1-3/8"
                                                      1-3/8"
                                                      1-3/8"
               Total Compressors Capacity
               Rack Capacity to Load Ratio
235.2
266.5
113%
               157%
[B-io;
                                                             Final Report, September 2, 2008
       www.epa.gov/gree n chill
                                                           Chillin' for the environment

-------
Table 3: DX Baseline MT Unit #3, +24°F, Refrigerant HFC-404A
DX Header Loads
# Loads

1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
Total Load
ID

SC1
SC2
SP
62
61
47
48
42
44
22
23
24
25
35
34
Load Description
Unit #3 Circuit Manifold
Rack#1 Subcooling
Rack #2 Subcooling
SPARE
AH-4, R=1-1/8"
AH-1 , AH-2A, AH-2B, R=1-3/8"
32' Produce Cases, R=1-1/8
32' Produce Cases, R=1-1/8
Seafood Room Coil, R=1-1/8
8' Salad Case, R=5/8"
36' Beverage Cases, R=1-3/8
36' Dairy Cases, R=1-3/8
24' Dairy Cases, R=1-3/8
24' Dairy Cases, R=1-3/8
Market Room Coil, R=1-1/8
Market Room Coil, R=1-1/8
#3 +24°F
DX Compr. Rack





Pos.
1
2
3
4
#3, Design conditions +24/1 10°F,
Compr. Model




Model MBTU
Remote 468.0
31.0
36.0

9.0
30.0
46.4
46.4
36.7
11.7
52.4
54.0
36.0
36.0
36.7
36.7
MBTU 468.0
Evap,°F
26
35
35

44
44
26
26
27
26
27
26
26
26
27
27

Load Line Sizes
Run
50
59
46

360
250
163
210
91
120
280
235
230
210
55
74

Supply
2-1/8"
1/2"
1/2"
None
1/2"
5/8"
7/8"
7/8"
5/8"
1/2"
7/8"
7/8"
7/8"
5/8"
1/2"
5/8"

Return
3-1/8"
1-1/8"
1-1/8"
None
7/8"
1-1/8"
1-5/8"
1-5/8"
1-3/8"
7/8"
1-5/8"
2-1/8"
1-5/8"
1-5/8"
1-1/8"
1-3/8"

Ctrl.Valves
Suction

ORIT-PI-311
ORIT-PI-413
Ball Valve
ORIT-PI-29
ORIT-PI-311
CDST-9-9
CDST-9-9
CDST-9-9
CDST-9-7
CDST-9-1 1
CDST-9-1 1
CDST-9-1 1
CDST-9-1 1
CDST-9-9
CDST-9-9

Subcooled liquid temp. =ambient temp.+10°F
Capacity, MBTU % Cap.
126.6 22%
126.6 22%
140.2 24%
189.9 33%
Rej.MBTU
172.4
172.4
190.9
257.7
Rej. %
1 36%
1 36%
1 36%
1 36%















                Total Compressors Capacity
                Rack Capacity to Load Ratio
583.3
793.4
136%
                125%
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Table 4: DX Baseline MT Unit #4, +20°F/15°F, Refrigerant HFC-404A
DX Header Loads +20°F
# Loads ID

1 SP
2 39
3 41
4 46
5
6 49A
7 49B
8 49C
9 45
10 43
11 55
12 60
13 56
14 51
15 57
16 58
17 28
18 27
19 31
20 32
21 33
Load Description Model
Unit #4 Circuit Manifold A Remote
• SPARE
• 12'x20'x10' Meat Cooler, R=7/8"
• 5'x8'x1 0' Seafood Cooler, R=5/8"
• 10'x24'x10' Produce Cir, R=7/8
• Loop 49, DR=7/8" & 5/8"
•• 8'x12'x10' Deli Cooler
•• 12'x12'x10' Deli Cooler
•• 8'x8'x10' Bakery Cooler
• 20' RL Produce Cases, R=1-1/8"
• 16' Produce Cases, R=7/8"
• 13' Floral Cases, R=7/8"
• 8' Deli Case, R=7/8"
• 32' Deli Island Cases, R=1-1/8
• 20' Deli Cases, R=1/2"
• 24' Deli Island Cases, R=1-1/8
• 32' Deli Island Cases, R=1-1/8
12'x38'x10' Dairy Cir, R=7/8"
36' RL Dairy Cases, R=1-3/8
20' Special Meat Case, R=7/8"
12'x128'x10' Chicken Cir, R=7/8"
24' Special Meat Cases, R=1-1/8"
Total Load #4 +20°F MBTU
MBTU
433.1

18.9
7.5
18.1
23.0
7.5
9.9
5.6
30.0
16.5
24.1
14.2
33.0
5.8
24.7
33.0
24.2
54.1
31.4
13.9
37.7
433.1
Evap,°F
22

22
22
22
22
22
22
22
22
22
22
22
22
22
22
22
22
22
22
22
22

DX Header Loads +15°F
# Loads ID

1 SP
2 38
3 37
4 36
Load Description Model
Unit #4 Circuit Manifold B Remote
• SPARE
• 12' Meat Cases, R=5/8"
• 24' Meat Cases, R=1-1/8"
• 24' Meat Cases, R=1-1/8"
Total Load #4 +15°F MBTU
[B-12]
MBTU
77.5


35.4
35.4
77.5
Evap,°F
17

17
17
17

Load
Run
50

55
53
140
200
25
25
25
140
110
360
240
270
245
230
160
120
117
93
57
65

Load
Run
50

90
60
37

Line Sizes
Supply
1-5/8"
None
5/8"
3/8"
1/2"
1/2"
3/8"
3/8"
3/8"
5/8"
1/2"
5/8"
1/2"
7/8"
3/8"
5/8"
5/8"
5/8"
7/8"
5/8"
5/8"
5/8"

Line Sizes
Supply
7/8"
None
5/8"
5/8"
5/8"


Return
3-1/8"
None
1-1/8"
7/8"
1-1/8"
1-1/8"
7/8"
7/8"
7/8"
1-3/8"
1-1/8"
1-3/8"
1-1/8"
1-5/8"
7/8"
1-3/8"
1-5/8"
1-3/8"
2-1/8"
1-3/8"
1-1/8"
1-3/8"


Return
1-5/8"
None
7/8"
1-3/8"
1-3/8"

Ctrl.Valves
Suction

Ball Valve
CDST-9-7
CDST-9-7
CDST-9-7
ORIT-PI-311



CDST-9-9
CDST-9-7
CDST-9-7
CDST-9-7
CDST-9-9
CDST-9-7
CDST-9-9
CDST-9-9
CDST-9-7
CDST-9-1 1
CDST-9-7
CDST-9-7
CDST-9-9

Ctrl.Valves
Suction

Ball Valve
CDS-9-9
CDS-9-9
CDS-9-9

Final Report, September 2, 2008
www.epa.gov/greenchill |
Chiltin'
for the environment

-------
Table 4: DX Baseline MT Unit #4, +20°F/15°F, Refrigerant HFC-404A (cont'd)
DX Compr. Rack #4, Design conditions +20/110°F, Subcooled liquid temp. =ambient temp.+10°F
         Pos.
         1
         2
         3
         4
Compr. Model
Capacity, MBTU
94.0
116.3
128.8
135.9
% Cap.
20%
24%
27%
29%
Rej.MBTU Rej. %
129.4     138%
160.7     138%
178      138%
186.7     137%
               Total Compressors Capacity
               Rack Capacity to Load Ratio
                              475.0
                                            110%
                         654.8
                    138%
DX Compr. Rack #4, Design conditions +15/110°F, Subcooled liquid temp. =ambienttemp.+10°F
         Pos.  Compr. Model                   Capacity, MBTU  % Cap.     Rej.MBTU Rej. %
         5                                   34.6           36%       48.5      140%
         6                                   62.0           64%       86.1      139%
               Total Compressors Capacity
               Rack Capacity to Load Ratio
                              96.6
                         134.6
                    139%
                                            125%
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-------
Table 5: DS-1, -25°F, Refrigerant HFC-404A
DS-1 Header Loads -25°F
# Loads  ID    Load Description                Model
4
5
6
7
8
9
10
11
6     12'+(1)EFz Island Case
5     12' Frozen Island Case
4     12'+(1)EFz Island Case
21    10 Drs Ice Cream Cases
20    10 Drs Ice Cream Cases
19    10 Drs Ice Cream Cases
18    5 Drs Ice Cream Cases
29    16'x24'x10'1C Freezer
30    12'x18'x10' Meet  Freezer
MBTU
10.5
7.6
10.5
14.1
14.1
14.1
7.0
20.8
13.6
Evap,°F
-12
-12
-12
-20
-20
                                                                          Load Line Sizes
                                                                         Run      Supply    Return
-20
-20
Total Load DS-1, -25°F
                                     MBTU
112.3
[B-14;
                                                                                                 Final Report, September 2, 2008
       www.epa.gov/gree n chill
                                                                                               Chillin' for the environment

-------
Table 6: DS-2, -14°F, Refrigerant HFC-404A
DS-2 Header Loads-14°F
# Loads ID    Load Description              Model
4        10    10 Drs Frozen Fd Cases
5        11    10 Drs Frozen Fd Cases
6        9     10 Drs Frozen Fd Cases
7        8     10 Drs Frozen Fd Cases
8        12    10 Drs Frozen Fd Cases
9        13    10 Drs Frozen Fd Cases
         7     10 Drs Frozen Fd Cases
         17    5 Drs Frozen Fd Cases
         16    10 Drs Frozen Fd Cases
         15    10 Drs Frozen Fd Cases
         14    10 Drs Frozen Fd Cases
i
ii
                                                                                 Load Line Sizes
MBTU
13.5
13.5
13.5
13.5
13.5
13.5
13.5
6.7
13.5
13.5
13.5
Evap,°F
-11
-11
-11
-11
-11
-11
-11
-11
-11
-11
-11
Run
294
264
268
237
233
202
205
302
289
257
226
Supply
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
5/8"
Return
1-3/8"
1-3/8"
1-3/8"
1-3/8"
1-3/8"
1-3/8"
1-3/8"
7/8"
1-3/8"
1-3/8"
1-3/8"
Total Load DS-2-14°F
                                             MBTU
141.7
Table 7: DS-3, +24°F, Refrigerant HFC-404A
DS-3 Header Loads +24°F
               Load Description               Model
               Rack#1 Subcooling
               Rack #2 Subcooling
               AH-1, AH-2A, AH-2B
               36' Beverage Cases
               36' Dairy Cases
               24' Dairy Cases
               24' Dairy Cases
# Loads
1
2
5
10
11
12
13
ID
SC1
SC2
61
22
23
24
25
Total Load DS-3, +24°F
                                             MBTU
 MBTU

 34.0
 30.0
 52.4
 54.0
 36.0
 36.0

 269.4
                                                                       Evap,°F
 Load Line Sizes
Run      Supply
Return
Ctrl.Valves
Suction
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
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                                                                                                      Chillin' for the environment

-------
Table 8: Distributed system, Loads DS-4a +20°F and DS-4b +15°F, Refrigerant HFC-404A
DS-4a, Header Loads +20°F
# Loads ID Load Description Model
Unit #4 Circuit Manifold A Remote
1 SP • SPARE
2 39 • 12'x20'x10' Meat Cooler, R=7/8"
16 58 • 32' Deli Island Cases, R=1-1/8
17 28 1 2'x38'x1 0' Dairy Cir, R=7/8"
18 27 36' RL Dairy Cases, R=1 -3/8
19 31 20' Special Meat Case, R=7/8"
20 32 1 2'x128'x10' Chicken Cir, R=7/8"
21 33 24' Special Meat Cases, R=1-1/8"
Total Load DS-4a, +20°F MBTU
DS-4b, Header Loads +15°F
# Loads ID Load Description Model
Unit #4 Circuit Manifold B Remote
1 SP • SPARE
2 38 • 12' Meat Cases, R=5/8"
3 37 • 24' Meat Cases, R=1 -1 /8"
4 36 • 24' Meat Cases, R=1 -1 /8"
Total Load DS-4b, +15°F MBTU
Table 9: DS-5, +24°F, Refrigerant HFC-404A
DS-5 Header Loads +24°F
# Loads ID Load Description Model
6 47 32' Produce Cases
7 48 32' Produce Cases
8 42 Seafood Room Coil
9 44 8' Salad Case
14 35 Market Room Coil
15 34 Market Room Coil
[B-16]
www.epa.gov/greenchill |
Load Line Sizes Ctrl.Valves
MBTU Evap,°F Run Supply Return Suction
213.2 22

18.9 22
33.0 22
24.2 22
54.1 22
31.4 22
13.9 22
37.7 22
213.2
Load Line Sizes Ctrl.Valves
MBTU Evap,°F Run Supply Return Suction
77.5 17


35.4 17
_£. . ,7
77.5

Load Line Sizes Ctrl.Valves
MBTU Evap,°F Run Supply Return Suction
46.4 26
46.4 26
36.7 27
11.7 26
367 27
Final Report, September 2, 2008
Chillin' for the environment

-------
Total Load DS-5, +24°F
                                                MBTU
            214.6
Table 10: DS-6a +20°F AND DS-6b -20°F, Refrigerant HFC-404A
DS-6a, Header
# Loads  ID
         SC3
               Loads +20°F
                Load Description
                System DS-6b Subcooling
          41     • 5'x8'x10' Seafood Cooler
          46     • 10'x24'x10' Produce Cir
          49A   •• 8'x12'x10' Deli Cooler
          49B   "12'x12'x10'Deli Cooler
          49C   •• 8'x8'x10' Bakery Cooler
          45     • 20' RL Produce Cases
          43     • 16'Produce Cases
          55     • 13'Floral Cases
          60     • 8' Deli Case
          56     • 32' Deli  Island Cases
          51     • 20' Deli  Cases
          57     • 24' Deli  Island Cases
3*        58     • 32' Deli  Island Cases
4*        62     AH-4, R=1-1/8"

Total Load DS-6a, +20°F
Model
MBTU
6.1
  Load Line Sizes            Ctrl.Valves
Run      Supply   Return    Suction
                                               MBTU
               245.0
DS-6b, Header Loads -20°F
# Loads   ID     Load Description
                Unit #4 Circuit Manifold B
1         SP    Spare
2*        54     8'x12'x10'Bakery Frz
3*        52     10'x12'x10'Bakery/Deli Frzr
2*        40     3 Drs Frozen Fd Cases
3*        59     3 Drs Frozen Fd Cases

Total Load DS-6b, -20°F
                                               Model
                                               Remote
                                               MBTU
               MBTU
               25.3
               25.3
           Evap,°F
           17
  Load Line Sizes
Run      Supply
Return
Ctrl.Valves
Suction
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
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-------
Table 11: Ambient Dry-Bulb Temperature in Atlanta, GA

Mid-pts
97.5
92.5
87.5
82.5
77.5
72.5
67.5
62.5
57.5
52.5
47.5
42.5
37.5
32.5
27.5
22.5
17.5
12.5

DB(F)
95 to 100
90 to 95
85 to 90
80 to 85
75 to 80
70 to 75
65 to 70
60 to 65
55 to 60
50 to 55
45 to 50
40 to 45
35 to 40
30 to 35
25 to 30
20 to 25
15 to 20
10 to 15
Total
Mrs
9
56
196
758
768
1314
885
1027
790
673
641
436
560
323
181
72
64
7
January February March April
Mrs






4
30
33
89
118
99
151
102
68
22
28

Mrs





7
23
73
78
75
95
50
84
70
45
36
29
7
Mrs




8
31
45
105
150
131
121
72
58
23




Mrs


2
24
59
82
93
157
106
81
53
30
31
2




May
Mrs


10
93
117
146
143
156
69
10








June
Mrs

18
40
154
139
222
110
35
2









July
Mrs
9
27
83
150
154
251
51
15
4









August September October November Dec
Mrs

10
56
182
142
247
84
23










Mrs


5
141
93
232
172
68
9









Mrs

1

14
56
84
108
190
120
78
64
25
4





Mrs





11
41
104
150
109
113
55
84
41
12



Mrs





1
11
71
69
100
77
105
148
85
56
14
7

[B-18;
Final Report, September 2, 2008


-------
Table 12: Ambient Dry-Bulb Temperature in Boulder, CO

Mid-pts
97.5
92.5
87.5
82.5
77.5
72.5
67.5
62.5
57.5
52.5
47.5
42.5
37.5
32.5
27.5
22.5
17.5
12.5
7.5
2.5
-2.5
-7.5

DB(F)
95 to 100
90 to 95
85 to 90
80 to 85
75 to 80
70 to 75
65 to 70
60 to 65
55 to 60
50 to 55
45 to 50
40 to 45
35 to 40
30 to 35
25 to 30
20 to 25
15 to 20
10to15
5 to 10
Oto5
-5 toO
-10 to -5
Total
Mrs
22
96
115
382
440
489
503
907
698
754
762
633
834
717
611
251
201
130
89
83
28
15
January
Mrs






6
17
16
44
63
67
62
102
113
65
58
60
27
20
11
13
February
Mrs







11
21
38
61
45
118
115
124
53
28
18
23
14
3

March
Mrs



4
1
14
10
28
41
55
69
80
114
135
95
37
36
18
7



April
Mrs




17
30
34
72
77
98
77
102
121
58
25
9






May
Mrs


2
26
52
55
51
87
110
120
114
59
52
16








June
Mrs
1
8
16
81
66
81
71
142
94
93
56
11










July
Mrs
18
45
37
108
114
112
119
178
11
2












August
Mrs
3
33
40
83
98
83
92
194
107
11












September
Mrs

10
20
66
55
59
73
99
103
111
79
22
22
1








October
Mrs



14
37
49
34
66
72
92
116
92
102
55
15







November
Mrs





6
13
13
36
60
63
92
143
114
94
42
44





December
Mrs








10
30
64
63
100
121
145
45
35
34
32
49
14
2
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
[B-19]


-------
Table 13: Ambient Dry-Bulb Temperature in Philadelphia

Mid-pts
97.5
92.5
87.5
82.5
77.5
72.5
67.5
62.5
57.5
52.5
47.5
42.5
37.5
32.5
27.5
22.5
17.5
12.5

DB(F)
95 to 100
90 to 95
85 to 90
80 to 85
75 to 80
70 to 75
65 to 70
60 to 65
55 to 60
50 to 55
45 to 50
40 to 45
35 to 40
30 to 35
25 to 30
20 to 25
15 to 20
10 to 15
Total
Mrs
3
52
104
477
656
907
619
983
625
540
576
552
1067
685
442
248
184
40
January February March
Mrs










21
86
142
119
153
101
98
24
Mrs









19
22
38
197
155
73
92
60
16
Mrs







21
53
66
122
113
196
105
56
12


April
Mrs



6
2
24
23
72
118
115
187
97
61
15




May
Mrs



13
68
100
96
203
123
89
36
15
1





June
Mrs

11
34
86
97
161
117
179
29
5
1







July
Mrs
3
30
53
184
198
161
62
52
1









August September October November Dec
Mrs

11
15
132
168
198
137
79
4









Mrs


2
52
96
200
96
165
91
13
5







Mrs



4
17
52
78
145
102
127
80
51
65
17
6



Mrs




10
11
10
66
96
93
66
75
148
106
35
4


Mrs







1
8
13
36
77
257
168
119
39
26

[ B-20 ;
Final Report, September 2, 2008


-------
 Figure 1a: Piping schematics of the baseline and alternative systems: Baseline and Alternative C

          BASELINE SYSTEM, ALTERNATE C (DISTRIBUTED)

                  MEDIUM-TEMPERATURE SYSTEMS                        LOW-TEMPERATURE SYSTEMS
                                              COMPONENTS AT REFRIGERATION SYSTEM
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                      Q
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                             LT LOADS
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
       www.epa.gov/gree n chill
                                                                              [ B-21 ]
                                                      Chillin' for the environment

-------
 Figure 1b: Piping schematics of the baseline and alternative systems: Alternative A
[ B-22 ;
ALTERNATE A [MTSC, LTCX)


 	MEWUN'VTEMFERATJJRE SY|TE_M5	
 I                                -ict.TC^.CKTOATime:


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                                                                                          Final Report, September 2, 2008
      www.epa.gov/gree n chill
                                                                                        Chillin' for the environment

-------
Figure 1c: Piping schematics of the baseline and alternative systems: Alternative B
i
^L~ERNATE B (MTSC. LTSCj
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Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
       www.epa.gov/gree n chill
                          [ B-23 ]
Chillin' for the environment

-------
 Figure 2: Fixture plan DX baseline
                                      DX  BASELINE  LAYOUT

                                            REFRIGERATION
                                              MECHANICAL


;



I

CENTER
/


	 i
L
                                              DX BASELINE
                                      DX-1,  SST-25°F: 129.6 kBTU/hr
                                      DX-2,  SST-14°F: 149.7 kBTU/hr.
                                      DX-3,  SST +24°F: 468.0 kBTU/hr
                                      DX-4a, SST +20°F: 433.1 kBTU/hr.
                                      DX-4b, SST +15°F:  77.5 kBTU/hr
                                                                                      Rack#1 Subcooling
                                                                                      Rack #2 Subcooling
AH-1, AH-2A, AH-2B
                                                                                      32' Produce Cases
                                                                                      32' Produce Cases
                                                                                      Seafood Room Coil
                                                                                      8' Salad Case
                                                                                      36' Beverage Cases
                                                                                      36' Dairy Cases
                                                                                      24' Dairy Cases
                                                                                      24' Dairy Cases
                                                                                      Market Room Coil
Market Room Coil
8'x12'x10' Bakery Frzr
10'x12'x10' Bakery/Deli Fr
                                                                                      12'+(1)EFz Island Case
                                                                                      12' Frozen Island Case
                                                                                      12'+(1)EFz Island Case
                                                                                      10 Drs Ice Cream Cases
                                                                                      10 Drs Ice Cream Cases
                                                                                      10 Drs Ice Cream Cases
                                                                                      5 Drs Ice Cream Cases
                                                                                      16'x24'x10' 1C Freezer
                                                                                      12'x18'x10' Meet Freezer
                                                                                                      31.0
                                                                                                      36.0
                                                                                                      9.0
                30.0
                                                                                                      46.4
                                                                                                      46.4
                                                                                                      36.7
                                                                                                      11.7
                                                                                                      52.4
                                                                                                      54.0
                                                                                                      36.0
                                                                                                      36.0
                                                                                                      36.7
                36.7
                8.0
                9.3
                                                                                                      10.5
                                                                                                      7.6
                                                                                                      10.5
                                                                                                      14.1
                                                                                                      14.1
                                                                                                      14.1
                                                                                                      7.0
                                                                                                      20.8
                                                                                                      13.6
                                                                                                           35
                                                                                                           35
                                                                                                           44
                     44
                                                                                                           26
                                                                                                           26
                                                                                                           27
                                                                                                           26
                                                                                                           27
                                                                                                           26
                                                                                                           26
                                                                                                           26
                                                                                                           27
                     27
                     -18
                     -18
                                                                                                           -12
                                                                                                           -12
                                                                                                           -12
                                                                                                           -20
                                                                                                           -20
                                                                                                           -20
                                                                                                           -20
                                                                                                           -22
                                                                                                           -18
[ B-24 ;
       www.epa.gov/greenchill
                           Final Report, September 2, 2008
                        Chiltin' for the environment

-------
 Figure 3: Fixture plan Distributed System
                                                                           jf    D        D
                                     [•'•  T- ---fii '"S^i
                                     •  • ••• • ••  -
                                        DISTRIBUTED
                                          SYSTEMS
                                  DS-1, SST -24°F:  112.3 kBTU/hr.
                                  DS-2, SST -13°F:  141.7 kBTU/hr.
                                  DS-3, SST +24°F:  269.4 kBTU/hr.
                                         5T +20°I       xBTU/hr.
                                  DS-4b, SST +15°F:  77.5 kBTU/hr.
                                  DS-5, SST + 24°F: 214.6 kBTU/hr.
                                  DS-6a, SST +20°F: 245.0 kBTU/hr.
                                  DS-6b, SST -20°F: 25.3 kBTU/hr.
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
      www.epa.gov/greenchill
                      [ B-25 ]
Chiltin' for the environment

-------
Appendix C


 Results Tables:
 Annual Energy Consumption, Power Input, and
 Weather Data, by Bin and Geographic Location
www.epa.gov/greenchill
Chiltin' for the environment

-------
Appendix C provides a detailed set of results tables. Presented for each baseline/alternative within each
geographical location, these tables present annual energy consumption, power input, and weather data by
bin. As illustrated in the tables, annual energy consumption per bin (kWh) is calculated by multiplying
power input per bin (kW) times the number of hours at the average ambient temperature for that bin.

Table C. 1:  Baseline (DX): bin and annual energy consumption for Atlanta, GA (DX)
Table C.2:  Baseline bin and annual energy consumption for Boulder, CO (DX)
Table C.3:  Baseline bin and annual energy consumption for Philadelphia, PA (DX)

Table C.4:  Alternative A bin and annual energy consumption for Atlanta, GA (MTS)
Table C.5:  Alternative A bin and annual energy consumption for Boulder, CO (MTS)
Table C.6:  Alternative A bin and annual energy consumption for Philadelphia, PA (MTS)

Table C.7:  Alternative B bin and annual energy consumption for Atlanta, GA (SC 50°F)
Table C.8:  Alternative B bin and annual energy consumption for Boulder, CO (SC 50°F)
Table C.9:  Alternative B bin and annual energy consumption for Philadelphia, PA (SC 50°F)

Table C. 10: Alternative B bin and annual energy consumption for Atlanta, GA (SC 40°F)
Table C. 11: Alternative B bin and annual energy consumption for Boulder, CO (SC 40°F)
Table C.12: Alternative B bin and annual energy consumption for Philadelphia, PA (SC 40°F)

Table C. 13: Alternative B bin and annual energy consumption for Atlanta, GA (SC 30°F)
Table C.14: Alternative B bin and annual energy consumption for Boulder, CO (SC 30°F)
Table C.15: Alternative B bin and annual energy consumption for Philadelphia, PA (SC 30°F)

Table C. 16: Alternative C bin and annual energy consumption for Atlanta, GA (DS)
Table C.17: Alternative C bin and annual energy consumption for Boulder, CO (DS)
Table C.I8: Alternative C bin and annual energy consumption for Philadelphia, PA (DS)
[C-2]
Final Report, September 2, 2008


-------
Table C.1 : Baseline bin and annual energy for Atlanta, GA (DX)
Amb.
Temp.
°F


95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond. LTSyst.
Temp.
°F


110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
Power
Input
kW

47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
, kWh)
MT
System
Power
Input
kW
124.3
112.5
102.24
92.31
84.09
76.64
68.75
62.27
56.35
55.79
55.24
54.69
54.15
54.15
54.15
54.15
54.15
54.15
54.15
54.15

Total
System
Power
kW

171.3
158.3
146.7
135.4
126.1
117.0
107.9
100.1
92.8
92.2
91.6
91.1
90.6
90.6
90.6
90.6
90.6
90.6
90.6
90.6

Weather
Bin Data
Atlanta, GA
h

9
56
196
758
768
1314
885
1027
790
673
641
436
560
323
181
72
64
7
0
0
8760
LT
System
Bin Energy
kWh

423
2562
8724
32644
32255
53077
34638
38867
28766
24505
23340
15876
20391
11761
6591
2622
2330
255
0
0
339,627
MT
System
Bin Energy
kWh

1118
6301
20039
69972
64578
100703
60847
63955
44517
37547
35407
23845
30323
17490
9801
3899
3465
379
0
0
594,186
Total
System
Bin Energy
kWh

1,542
8,864
28,763
102,617
96,834
153,780
95,485
102,822
73,282
62,052
58,747
39,720
50,714
29,251
16,391
6,520
5,796
634
0
0
933,813
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
[C-3]


-------
Table C.2 : Baseline Total Bin and Annual Energy for Boulder, CO (DX)
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond. LTSyst.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
Power
Input
kW
47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
, kWh)
MT System
Power
Input
kW
124.3
112.5
102.24
92.31
84.09
76.64
68.75
62.27
56.35
55.79
55.24
54.69
54.15
54.15
54.15
54.15
54.15
54.15
54.15
54.15

Total
System
Power
kW
171.3
158.3
146.7
135.4
126.1
117.0
107.9
100.1
92.8
92.2
91.6
91.1
90.6
90.6
90.6
90.6
90.6
90.6
90.6
90.6

Weather
Bin Data
Boulder, CO
h
22
96
115
382
440
489
503
907
698
754
762
633
834
717
611
251
201
130
89
126
8760
LT
System
Bin Energy
kWh
1035
4392
5119
16451
18480
19752
19687
34326
25416
27455
27746
23049
30368
26108
22248
9139
7319
4734
3241
4588
330,651
MT
System
Bin Energy
kWh
2734
10802
11757
35263
36998
37476
34583
56482
39332
42066
42091
34619
45159
38824
33084
13591
10884
7039
4819
6823
544,427
Total
System
Bin Energy
kWh
3,769
15,195
16,876
51,714
55,478
57,229
54,270
90,808
64,748
69,521
69,837
57,667
75,527
64,932
55,332
22,731
18,203
11,773
8,060
11,411
875,078
[C-4]
Final Report, September 2, 2008


-------
Table C.3: Baseline bin and annual energy for Philadelphia, PA
Amb.
Temp.
°F


95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F


110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Syst.
Power
Input
kW

47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
kWh)
MT
System
Power
Input
kW
124.3
112.5
102.24
92.31
84.09
76.64
68.75
62.27
56.35
55.79
55.24
54.69
54.15
54.15
54.15
54.15
54.15
54.15
54.15
54.15

Total
System
Power
kW

171.3
158.3
146.7
135.4
126.1
117.0
107.9
100.1
92.8
92.2
91.6
91.1
90.6
90.6
90.6
90.6
90.6
90.6
90.6
90.6

Weather
Bin Data
Philadelphia,
PA
h
3
52
104
477
656
907
619
983
625
540
576
552
1067
685
442
248
184
40


8760
(DX)
LT
System
Bin
Energy
kWh
141
2379
4629
20543
27551
36637
24227
37202
22758
19663
20973
20100
38852
24942
16094
9030
6700
1456
0
0
333,877

MT
System
Bin Energy
kWh

373
5851
10633
44033
55161
69511
42559
61215
35219
30127
31816
30189
57776
37091
23933
13429
9963
2166
0
0
561,044

Total
System
Bin
Energy
kWh
514
8,230
15,262
64,575
82,712
106,148
66,786
98,417
57,976
49,789
52,790
50,288
96,628
62,034
40,028
22,459
16,663
3,622
0
0
894,921
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
[C-5]


-------
Table C.4: Alternative A
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond. LT Syst.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
Power
Input
kW
47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
, kWh)
bin and annual energy for Atlanta,
MT Syst.
Power
Input
kW
132.7
121.5
109.7
100.8
91.5
84.4
76.8
69.9
64.1
58.1
52.7
47.8
43.3
43.3
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
179.7
167.2
154.2
143.9
133.5
124.7
115.9
107.8
100.5
94.5
89.1
84.2
79.7
79.7
79.7
79.7
79.7
79.7
79.7
79.7

Weather
Bin Data
Atlanta, GA
h
9
56
196
758
768
1314
885
1027
790
673
641
436
560
323
181
72
64
7
0
0
8760
GA (MTS)
LT
System
Bin Energy
kWh
423
2,562
8,724
32,644
32,255
53,077
34,638
38,867
28,766
24,505
23,340
15,876
20,391
11,761
6,591
2,622
2,330
255
0
0
339,627

MT
System
Bin Energy
kWh
1,194
6,803
21,508
76,439
70,241
110,841
67,943
71,796
50,622
39,117
33,798
20,842
24,252
13,988
7,839
3,118
2,772
303
0
0
623,416

Total
System
Bin Energy
kWh
1,618
9,365
30,232
109,083
102,497
163,918
102,581
110,663
79,388
63,623
57,138
36,718
44,643
25,749
14,429
5,740
5,102
558
0
0
963,043
[C-6]
Final Report, September 2, 2008


-------
Table C.S: Alternative A bin and annual energy for Boulder,
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond. LT Syst.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
Power
Input
kW
47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
, kWh)
MT Syst.
Power
Input
kW
132.7
121.5
109.7
100.8
91.5
84.4
76.8
69.9
64.1
58.1
52.7
47.8
43.3
43.3
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
179.7
167.2
154.2
143.9
133.5
124.7
115.9
107.8
100.5
94.5
89.1
84.2
79.7
79.7
79.7
79.7
79.7
79.7
79.7
79.7

Weather
Bin Data
Boulder, CO
h
22
96
115
382
440
489
503
907
698
754
762
633
834
717
611
251
201
130
89
126
8760
CO (MTS)
LT
System
Bin Energy
kWh
1,035
4,392
5,119
16,451
18,480
19,752
19,687
34,326
25,416
27,455
27,746
23,049
30,368
26,108
22,248
9,139
7,319
4,734
3,241
4,588
330,651

MT
System
Bin Energy
kWh
2,919
1 1 ,662
12,620
38,522
40,242
41,249
38,616
63,407
44,727
43,825
40,178
30,260
36,118
31,051
26,460
10,870
8,705
5,630
3,854
5,457
536,371

Total
System
Bin Energy
kWh
3,954
16,054
17,738
54,973
58,722
61,001
58,303
97,733
70,142
71,280
67,924
53,308
66,486
57,158
48,708
20,009
16,024
10,363
7,095
10,045
867,022
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
[C-7]


-------
Table C.6: Alternative A bin and annual energy for Philadelphia, PA (MTS)
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond. LTSyst.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
Power
Input
kW
47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
, kWh)
MT Syst.
Power
Input
kW
132.7
121.5
109.7
100.8
91.5
84.4
76.8
69.9
64.1
58.1
52.7
47.8
43.3
43.3
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
179.7
167.2
154.2
143.9
133.5
124.7
115.9
107.8
100.5
94.5
89.1
84.2
79.7
79.7
79.7
79.7
79.7
79.7
79.7
79.7

Weather
Bin Data
Philadelphia, PA
h
3
52
104
477
656
907
619
983
625
540
576
552
1067
685
442
248
184
40
0
0
8760
LT
System
Bin Energy
kWh
141
2,379
4,629
20,543
27,551
36,637
24,227
37,202
22,758
19,663
20,973
20,100
38,852
24,942
16,094
9,030
6,700
1,456
0
0
333,877
MT
System
Bin Energy
kWh
398
6,317
11,413
48,102
59,998
76,509
47,521
68,720
40,049
31,387
30,371
26,387
46,208
29,665
19,142
10,740
7,968
1,732
0
0
562,628
Total
System
Bin Energy
kWh
539
8,696
16,042
68,645
87,549
113,146
71,748
105,922
62,807
51,049
51,344
46,487
85,060
54,607
35,236
19,770
14,668
3,189
0
0
896,505
[C-8]
Final Report, September 2, 2008


-------
Table C.7: Alternative B
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond. LT Sys.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
Power
Input
kW
47.85
46.85
45.64
44.22
42.84
41.51
40.50
39.24
38.04
36.86
35.70
34.29
33.21
31.06
29.05
29.05
29.05
29.05
29.05
29.05
, kWh)
bin and annual energy for Atlanta, GA (SC 50°F)
MT Sys.
Power
Input
kW
133.1
121.8
110.0
101.1
91.6
84.5
76.9
70.0
64.1
58.2
52.8
47.8
43.3
43.3
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
180.9
168.7
155.6
145.3
134.5
126.0
117.4
109.2
102.2
95.0
88.5
82.1
76.5
74.4
72.4
72.4
72.4
72.4
72.4
72.4

Weather
Bin Data
Atlanta, GA
h
9
56
196
758
768
1314
885
1027
790
673
641
436
560
323
181
72
64
7
0
0
8760
LT
System
Bin Energy
kWh
431
2,624
8,945
33,517
32,902
54,542
35,846
40,295
30,052
24,809
22,886
14,950
18,598
10,031
5,257
2,091
1,859
203
0
0
339,838
MT
System
Bin Energy
kWh
1,198
6,821
21,560
76,605
70,377
111,028
68,040
71,882
50,670
39,145
33,814
20,847
24,252
13,988
7,839
3,118
2,772
303
0
0
624,258
Total
System
Bin Energy
kWh
1,628
9,445
30,505
110,121
103,279
165,569
103,886
112,177
80,722
63,954
56,700
35,797
42,850
24,019
13,096
5,209
4,631
506
0
0
964,096
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
;c-9]


-------
Table C.8: Alternative B total bin and annual energy for
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
LT Sys.
Power
Input
kW
47.85
46.85
45.64
44.22
42.84
41.51
40.50
39.24
38.04
36.86
35.70
34.29
33.21
31.06
29.05
29.05
29.05
29.05
29.05
29.05
, kWh)
MT Sys.
Power
Input
kW
133.1
121.8
110.0
101.1
91.6
84.5
76.9
70.0
64.1
58.2
52.8
47.8
43.3
43.3
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
180.9
168.7
155.6
145.3
134.5
126.0
117.4
109.2
102.2
95.0
88.5
82.1
76.5
74.4
72.4
72.4
72.4
72.4
72.4
72.4

Weather
Bin Data
Boulder, CO
h
22
96
115
382
440
489
503
907
698
754
762
633
834
717
611
251
201
130
89
126
8760
Boulder, CO (SC 50°F)
LT
System
Bin Energy
kWh
1,053
4,498
5,248
16,891
18,850
20,297
20,373
35,587
26,552
27,795
27,206
21,705
27,698
22,267
17,748
7,291
5,838
3,776
2,585
3,660
316,919
MT
System
Bin Energy
kWh
2,928
11,693
12,650
38,606
40,320
41,319
38,671
63,483
44,769
43,856
40,197
30,267
36,118
31,051
26,460
10,870
8,705
5,630
3,854
5,457
536,903
Total
System
Bin Energy
kWh
3,980
16,191
17,899
55,497
59,170
61,616
59,045
99,069
71,322
71,652
67,403
51,972
63,816
53,318
44,208
18,161
14,543
9,406
6,439
9,117
853,822
Final Report, September 2, 2008


-------
Table C.9: Alternative B 50° F Total Bin and Annual Energy for Philadelphia, PA (SC 50° F)
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
LT Sys.
Power
Input
kW
47.85
46.85
45.64
44.22
42.84
41.51
40.50
39.24
38.04
36.86
35.70
34.29
33.21
31.06
29.05
29.05
29.05
29.05
29.05
29.05
, kWh)
MT Sys.
Power
Input
kW
133.1
121.8
110.0
101.1
91.6
84.5
76.9
70.0
64.1
58.2
52.8
47.8
43.3
43.3
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
180.9
168.7
155.6
145.3
134.5
126.0
117.4
109.2
102.2
95.0
88.5
82.1
76.5
74.4
72.4
72.4
72.4
72.4
72.4
72.4

Weather
Bin Data
Philadelphia, PA
h
3
52
104
477
656
907
619
983
625
540
576
552
1067
685
442
248
184
40
0
0
8760
LT
System
Bin Energy
kWh
144
2,436
4,746
21,092
28,104
37,648
25,072
38,568
23,775
19,906
20,565
18,928
35,436
21,273
12,839
7,204
5,345
1,162
0
0
324,243
MT
System
Bin Energy
kWh
399
6,334
1 1 ,440
48,206
60,113
76,638
47,590
68,802
40,087
31,409
30,385
26,394
46,208
29,665
19,142
10,740
7,968
1,732
0
0
563,253
Total
System
Bin Energy
kWh
543
8,770
16,186
69,298
88,217
114,286
72,661
107,371
63,863
51,315
50,950
45,321
81,645
50,939
31,980
17,944
13,313
2,894
0
0
887,496
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
[C-11]


-------
Table C.10: Alternative B
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Sys.
Power
Input
kW
45.34
44.44
43.33
42.02
40.74
39.51
38.31
37.39
36.03
34.94
33.85
32.57
31.52
30.28
29.05
29.05
29.05
29.05
29.05
29.05
kWh)
(40°F) Total Bin and Annual Energy for Atlanta, GA (SC 40°F)
MT Sys.
Power
Input
kW
134.4
123.1
111.2
102.3
92.8
85.6
77.9
71.0
65.1
59.0
53.6
48.6
44.0
43.7
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
179.76
167.54
154.57
144.27
133.52
125.10
116.23
108.37
101.11
93.98
87.42
81.15
75.54
73.94
72.35
72.35
72.35
72.35
72.35
72.35

Weather
Bin Data
Atlanta, GA
h
9
56
196
758
768
1314
885
1027
790
673
641
436
560
323
181
72
64
7
0
0
8760
LT
System
Bin Energy
kWh
408
2,489
8,492
31,848
31,292
51,915
33,905
38,403
28,466
23,511
21,700
14,202
17,653
9,779
5,257
2,091
1,859
203
0
0
323,473
MT
System
Bin Energy
kWh
1,210
6,894
21,803
77,511
71,249
112,467
68,959
72,891
51,408
39,734
34,338
21,180
24,649
14,102
7,839
3,118
2,772
303
0
0
632,425
Total
System
Bin Energy
kWh
1,618
9,382
30,295
109,359
102,540
164,382
102,864
1 1 1 ,294
79,873
63,246
56,038
35,382
42,302
23,881
13,096
5,209
4,631
506
0
0
955,899
[C-12]
Final Report, September 2, 2008


-------
Table C.11: Alternative B
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Sys.
Power
Input
kW
45.34
44.44
43.33
42.02
40.74
39.51
38.31
37.39
36.03
34.94
33.85
32.57
31.52
30.28
29.05
29.05
29.05
29.05
29.05
29.05
kWh)
(40°F) Total Bin and Annual Energy for Boulder, CO (SC 40°
MT Sys.
Power
Input
kW
134.4
123.1
111.2
102.3
92.8
85.6
77.9
71.0
65.1
59.0
53.6
48.6
44.0
43.7
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
179.76
167.54
154.57
144.27
133.52
125.10
116.23
108.37
101.11
93.98
87.42
81.15
75.54
73.94
72.35
72.35
72.35
72.35
72.35
72.35

Weather
Bin Data
Boulder, CO
h
22
96
115
382
440
489
503
907
698
754
762
633
834
717
611
251
201
130
89
126
8760
LT
System
Bin Energy
kWh
998
4,266
4,983
16,050
17,927
19,320
19,270
33,916
25,151
26,341
25,796
20,619
26,290
21,708
17,748
7,291
5,838
3,776
2,585
3,660
303,532
MT
System
Bin Energy
kWh
2,957
11,818
12,793
39,062
40,820
41,854
39,194
64,374
45,421
44,516
40,820
30,750
36,709
31,304
26,460
10,870
8,705
5,630
3,854
5,457
543,368
F)
Total
System
Bin Energy
kWh
3,955
16,084
17,775
55,112
58,747
61,174
58,464
98,290
70,572
70,858
66,616
51,369
62,999
53,012
44,208
18,161
14,543
9,406
6,439
9,117
846,900
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
[C-13]


-------
Table C.12: Alternative
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Sys.
Power
Input
kW
45.34
44.44
43.33
42.02
40.74
39.51
38.31
37.39
36.03
34.94
33.85
32.57
31.52
30.28
29.05
29.05
29.05
29.05
29.05
29.05
kWh)
B 40°F Total Bin and Annual Energy for Philadelphia, PA (SC 40°F)
MT Sys.
Power
Input
kW
134.4
123.1
111.2
102.3
92.8
85.6
77.9
71.0
65.1
59.0
53.6
48.6
44.0
43.7
43.3
43.3
43.3
43.3
43.3
43.3

Total
System
Power
kW
179.76
167.54
154.57
144.27
133.52
125.10
116.23
108.37
101.11
93.98
87.42
81.15
75.54
73.94
72.35
72.35
72.35
72.35
72.35
72.35

Weather
Bin Data
Philadelphia, PA
h
3
52
104
477
656
907
619
983
625
540
576
552
1067
685
442
248
184
40


8760
LT
System
Bin Energy
kWh
136
2,311
4,506
20,042
26,728
35,835
23,714
36,758
22,520
18,865
19,499
17,981
33,635
20,739
12,839
7,204
5,345
1,162
0
0
309,817
MT
System
Bin Energy
kWh
403
6,401
11,569
48,777
60,858
77,631
48,232
69,768
40,671
31,882
30,856
26,815
46,965
29,907
19,142
10,740
7,968
1,732
0
0
570,318
Total
System
Bin Energy
kWh
539
8,712
16,075
68,818
87,587
113,466
71,946
106,526
63,191
50,747
50,356
44,796
80,600
50,646
31,980
17,944
13,313
2,894
0
0
880,135
[C-14]
Final Report, September 2, 2008


-------
Table C.13: Alternative B
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Sys.
Power
Input
kW
43.41
42.27
41.25
39.77
38.60
37.71
36.59
35.50
34.45
33.42
32.20
31.00
30.01
28.84
27.68
27.68
27.68
27.68
27.68
27.68
kWh)
30°F Total Bin and
MT Sys.
Power
Input
kW
135.60
124.25
112.33
103.31
93.78
86.56
78.84
71.84
65.90
59.81
54.29
49.25
44.64
44.30
43.97
43.97
43.97
43.97
43.97
43.97

Total
System
Power
kW
179.00
166.53
153.58
143.08
132.37
124.26
115.43
107.34
100.35
93.23
86.49
80.25
74.65
73.14
71.65
71.65
71.65
71.65
71.65
71.65

Annual Energy for Atlanta
Weather
Bin Data
Atlanta, GA
h
9
56
196
758
768
1314
885
1027
790
673
641
436
560
323
181
72
64
7
0
0
8760
LT
System
Bin Energy
kWh
391
2,367
8,085
30,144
29,644
49,545
32,381
36,460
27,217
22,489
20,639
13,514
16,805
9,314
5,010
1,993
1,772
194
0
0
307,964
, GA (SC 30°F)
MT
System
Bin Energy
kWh
1,220
6,958
22,017
78,312
72,020
113,739
69,771
73,782
52,060
40,255
34,802
21,474
25,000
14,310
7,958
3,166
2,814
308
0
0
639,967
Total
System
Bin Energy
kWh
1,611
9,325
30,102
108,456
101,664
163,283
102,152
110,242
79,277
62,744
55,441
34,988
41,805
23,625
12,968
5,159
4,586
502
0
0
947,931
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
[C-15]


-------
Table C.14: Alternative B 30°F Total Bin and
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Sys.
Power
Input
kW
43.41
42.27
41.25
39.77
38.60
37.71
36.59
35.50
34.45
33.42
32.20
31.00
30.01
28.84
27.68
27.68
27.68
27.68
27.68
27.68
kWh)
MT Sys.
Power
Input
kW
135.60
124.25
112.33
103.31
93.78
86.56
78.84
71.84
65.90
59.81
54.29
49.25
44.64
44.30
43.97
43.97
43.97
43.97
43.97
43.97

Total
System
Power
kW
179.00
166.53
153.58
143.08
132.37
124.26
115.43
107.34
100.35
93.23
86.49
80.25
74.65
73.14
71.65
71.65
71.65
71.65
71.65
71.65

Annual Energy for Boulder,
Weather
Bin Data
Boulder, CO
h
22
96
115
382
440
489
503
907
698
754
762
633
834
717
611
251
201
130
89
126
8760
LT
System
Bin Energy
kWh
955
4,058
4,744
15,191
16,983
18,438
18,404
32,200
24,048
25,196
24,534
19,620
25,028
20,676
16,912
6,948
5,564
3,598
2,464
3,488
289,049
CO (SC 30°F)
MT
System
Bin Energy
kWh
2,983
11,928
12,918
39,466
41,261
42,327
39,655
65,161
45,997
45,100
41,372
31,177
37,232
31,766
26,865
11,036
8,838
5,716
3,913
5,540
550,253
Total
System
Bin Energy
kWh
3,938
15,986
17,662
54,657
58,245
60,765
58,059
97,361
70,045
70,296
65,906
50,797
62,260
52,442
43,777
17,984
14,401
9,314
6,377
9,028
839,302
Final Report, September 2, 2008


-------
Table C.15: Alternative B bin and annual energy for Philadelphia, PA (SC SOT)
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
Annual (hours
LT Sys.
Power
Input
kW
43.41
42.27
41.25
39.77
38.60
37.71
36.59
35.50
34.45
33.42
32.20
31.00
30.01
28.84
27.68
27.68
27.68
27.68
27.68
27.68
, kWh)
MT Sys.
Power
Input
kW
135.60
124.25
112.33
103.31
93.78
86.56
78.84
71.84
65.90
59.81
54.29
49.25
44.64
44.30
43.97
43.97
43.97
43.97
43.97
43.97

Total
System
Power
kW
179.00
166.53
153.58
143.08
132.37
124.26
115.43
107.34
100.35
93.23
86.49
80.25
74.65
73.14
71.65
71.65
71.65
71.65
71.65
71.65

Weather
Bin Data
Philadelphia, PA
h
3
52
104
477
656
907
619
983
625
540
576
552
1067
685
442
248
184
40


8760
LT
System
Bin Energy
kWh
130
2,198
4,290
18,969
25,321
34,199
22,648
34,898
21,533
18,045
18,546
17,110
32,020
19,753
12,235
6,865
5,093
1,107
0
0
294,959
MT
System
Bin Energy
kWh
407
6,461
11,683
49,281
61,517
78,509
48,800
70,621
41,187
32,300
31,273
27,187
47,634
30,348
19,434
10,904
8,090
1,759
0
0
577,396
Total
System
Bin Energy
kWh
537
8,659
15,973
68,250
86,838
112,708
71,449
105,519
62,719
50,345
49,819
44,297
79,654
50,102
31,669
17,769
13,183
2,866
0
0
872,355
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies
[C-17]


-------
Table C.16: Alternative C Total Bin and Annual Energy for Atlanta, GA (DS)
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Sys.
Power
Input
kW
47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
kWh)
MT Sys.
Power
Input
kW
116.22
107.34
96.86
87.85
79.78
72.47
65.34
58.98
53.19
52.69
52.20
51.72
51.24
51.24
51.24
51.24
51.24
51.24
51.24
51.24

Total
System
Power
kW
163.26
153.09
141.37
130.92
121.77
112.86
104.48
96.82
89.60
89.11
88.62
88.13
87.65
87.65
87.65
87.65
87.65
87.65
87.65
87.65

Weather
Bin Data
Atlanta, GA
h
9
56
196
758
768
1314
885
1027
790
673
641
436
560
323
181
72
64
7
0
0
8760
LT
System
Bin Energy
kWh
423
2562
8724
32644
32255
53077
34638
38867
28766
24505
23340
15876
20391
11761
6591
2622
2330
255
0
0
339,627
MT
System
Bin Energy
kWh
1046
6011
18985
66592
61267
95228
57830
60572
42021
35463
33463
22549
28693
16550
9274
3689
3279
359
0
0
562,871
Total
System
Bin Energy
kWh
1,469
8,573
27,709
99,237
93,523
148,305
92,468
99,439
70,786
59,969
56,803
38,425
49,084
28,311
15,865
6,311
5,610
614
0
0
902,499
[c-18;
Final Report, September 2, 2008


-------
Table C.17: Alternative C Total Bin and Annual Energy for Boulder, CO (DS)
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Sys.
Power
Input
kW
47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
kWh)
MT Sys.
Power
Input
kW
116.22
107.34
96.86
87.85
79.78
72.47
65.34
58.98
53.19
52.69
52.20
51.72
51.24
51.24
51.24
51.24
51.24
51.24
51.24
51.24

Total
System
Power
kW
163.26
153.09
141.37
130.92
121.77
112.86
104.48
96.82
89.60
89.11
88.62
88.13
87.65
87.65
87.65
87.65
87.65
87.65
87.65
87.65

Weather
Bin Data
Boulder, CO
h
22
96
115
382
440
489
503
907
698
754
762
633
834
717
611
251
201
130
89
126
8,760
LT
System
Bin Energy
kWh
1035
4392
5119
16451
18480
19752
19687
34326
25416
27455
27746
23049
30368
26108
22248
9139
7319
4734
3241
4588
330,651
MT
System
Bin Energy
kWh
2557
10305
11139
33560
35101
35439
32868
53494
37127
39731
39779
32738
42733
36738
31306
12861
10299
6661
4560
6456
515,452
Total
System
Bin Energy
kWh
3592
14697
16258
50011
53581
55191
52555
87820
62543
67186
67525
55787
73100
62845
53554
22000
17618
11395
7801
11044
846,102
Theoretical Analysis of Alternative Supermarket Refrigeration Technologies


-------
Table C.18: Alternative C Total Bin and Annual Energy for Philadelphia,
Amb.
Temp.
°F

95-100
90-95
85-90
80-85
75-80
70-75
65-70
60-65
55-60
50-55
45-50
40-45
35-40
30-35
25-30
20-25
15-20
10-15
5-10
0-5
Annual
Cond.
Temp.
°F

110
105
100
95
90
85
80
75
70
65
60
55
50
50
50
50
50
50
50
50
(hours,
LT Sys.
Power
Input
kW
47.04
45.75
44.51
43.07
42.00
40.39
39.14
37.85
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
36.41
kWh)
MT Sys.
Power
Input
kW
116.22
107.34
96.86
87.85
79.78
72.47
65.34
58.98
53.19
52.69
52.20
51.72
51.24
51.24
51.24
51.24
51.24
51.24
51.24
51.24

Total
System
Power
kW
163.26
153.09
141.37
130.92
121.77
112.86
104.48
96.82
89.60
89.11
88.62
88.13
87.65
87.65
87.65
87.65
87.65
87.65
87.65
87.65

Weather
Bin Data
Philadelphia, PA
h
3
52
104
477
656
907
619
983
625
540
576
552
1067
685
442
248
184
40


8760
LT
System
Bin Energy
kWh
141
2379
4629
20543
27551
36637
24227
37202
22758
19663
20973
20100
38852
24942
16094
9030
6700
1456
0
0
333,877
PA (DS)
MT
System
Bin Energy
kWh
349
5582
10074
41906
52333
65732
40448
57977
33244
28455
30069
28549
54671
35098
22647
12707
9428
2050
0
0
531,317

Total
System
Bin Energy
kWh
490
7,961
14,703
62,448
79,884
102,368
64,675
95,179
56,002
48,118
51,043
48,648
93,523
60,040
38,741
21,737
16,128
3,506
0
0
865,194
[ c-20;
Final Report, September 2, 2008


-------
Acknowledgements

The GreenChill Advanced Refrigeration Partnership is an EPA cooperative alliance with the
supermarket industry and other stakeholders to promote advanced technologies, strategies, and
practices that reduce refrigerant charges and emissions of ozone-depleting substances and
greenhouse gases
Working with EPA, GreenChill Partners:

   •   Transition to non-ozone-depleting
       refrigerants;
   •   Reduce refrigerant charges;
   •   Reduce both ozone-depleting and
       greenhouse gas refrigerant emissions;
       and
   •   Promote supermarkets' adoption of
       advanced refrigeration technologies.
Lead author: Georgi Kazachki, Cryotherm.

Special thanks to others who contributed to the study: Julius Banks, Cynthia Gage, David S.
Godwin, and Bella Maranion, EPA; and Joanna L. Pratt, Stratus Consulting Inc.

Special thanks to members of the technical review panel who provided valuable input in
designing the study: Wayne Rosa, Food Lion, LLC; Harrison Horning, Hannaford and Sweetbay;
Chris LaPietra, Honeywell Refrigerants; Rob Fennell, Honeywell; Ron Vogl, Honeywell; Kathy
Loftus, Whole Foods Market; Stephen Sloan, Publix Super Markets, Inc; and Cliff Timko, Giant
Eagle, Inc.

Special thanks also to the peer reviewers of the report: Bernard Adebayo-Ige, Albertsons; and
Ken Welter, Stop and Shop.


-------