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Design & Engineering Services

Performance Comparison of Three High Efficiency Medium-Temperature Display Cases ET 06-07

Prepared by:

Design & Engineering Services Customer Service Business Unit Southern California Edison

June 22, 2009

Performance Comparison of Three High Efficiency Medium-Temperature Display Cases ET 06.07

Southern California Edison Design & Engineering Services March 2009

Acknowledgements

Southern California Edison’s Design & Engineering Services (D&ES) group is responsible for this project. It was developed as part of Southern California Edison’s Emerging Technologies program under internal project number ET 06.07. D&ES project manager Rafik Sarhadian in collaboration with Devin Rauss, Bruce Coburn, John Lutton, and Scott Mitchell conducted this technology evaluation with overall guidance and management from Paul Delaney, and Ramin Faramarzi. For more information on this project, contact [email protected].

Disclaimer

This report was prepared by Southern California Edison (SCE) and funded by California utility customers under the auspices of the California Public Utilities Commission. Reproduction or distribution of the whole or any part of the contents of this document without the express written permission of SCE is prohibited. This work was performed with reasonable care and in accordance with professional standards. However, neither SCE nor any entity performing the work pursuant to SCE’s authority make any warranty or representation, expressed or implied, with regard to this report, the merchantability or fitness for a particular purpose of the results of the work, or any analyses, or conclusions contained in this report. The results reflected in the work are generally representative of operating conditions; however, the results in any other situation may vary depending upon particular operating conditions.


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ABBREVIATIONS AND ACRONYMS A Surface Area, square-feet (ft2)

A/C Air Conditioning

AHU Air Handling Unit

ANN Artificial Neural Network

Cfm Cubic feet per minute

CTAC Customer Technology Application Center

DAG Discharge Air Grille

DAT Discharge Air Temperature

D&ES Design and Engineering Services

DB Dry-Bulb Temperature, oF

DC Direct Current

DX Direct Expansion

EFLH Equivalent Full Load Hours

EXV Electronic Expansion Valve

hp Horsepower

kW Kilowatt

kWh Kilowatt hour

LMTD Low-Mean Temperature Difference

LT Low Temperature

MT Medium Temperature

RAG Return Air Grille

RAT Return Air Temperature

RH Relative Humidity, %

RTTC Refrigeration and Thermal Test Center


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SET Saturated Evaporating Temperature

SCE Southern California Edison

SCT Saturated Condensing Temperature, oF

TD Temperature Difference, oF

TXV Thermostatic Expansion Valve

U Overall Heat Transfer Coefficient, Btu/hr-ft2-oF

VAV Variable Air Volume

VFD Variable Frequency Drive

VSD Variable Speed Drive

WB Wet-Bulb Temperature, oF

T Temperature Differential, oF


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FIGURES Figure 1. Percentage Breakdown of Display Cases by Type in a

Typical Supermarket [Ref 2] ......................................... 4 Figure 2. Schematics of a Typical Open Vertical Refrigerated

Display Case and Air Circulation Pattern (Side View) ......... 6 Figure 3. Refrigeration Load for for Typical Medium-Temperature

Open Vertical Refrigerated Display Case at 75oF Dry Bulb and 55% Relative Humidity [Ref 1] ......................... 7

Figure 4. Schematic Diagram of the Air Conditioning and Heating System of the RTTC’s Controlled Environment Room ........ 9

Figure 5. Custom Raised Frame Assembly and Special Drain Piping/Valve Arrangement .......................................... 11

Figure 6. Simulated and Dummy Products Used in the Display Case ........................................................................ 11

Figure 7. Location of Product Simulators Inside the Display Case ... 12 Figure 8. Location of Sensors for Open Vertical Multi-Deck Display

Cases ...................................................................... 14 Figure 9. High Precision Digital Scale Used to Measure the Weight

of condensate Collected .............................................. 15 Figure 10. Schematics of Inner and Outer Shell of the Case and

Insulation Between Them ........................................... 25 Figure 11. Surfaces Participating in Display Case Radiation Heat

Transfer ................................................................... 27 Figure 12. Photograph of Hill Phoenix’s 8-foot, 5-deck Display Case . 31 Figure 13. Schematic of the 8-foot, 5-Deck Display Hill Phoenix

Case (Courtesy of Hill Phoenix) ................................... 31 Figure 14. Photograph of Hussmann’s 8-foot, 4-deck Display Case .. 33 Figure 15. Schematic of the 8-foot, 4-Deck Hussmann Display

Case (Courtesy of Hussmann) ..................................... 33 Figure 16. Photograph of Tyler’s 8-foot, 5-deck Display Case .......... 35 Figure 17. Schematic of the 8-foot, 5-Deck Tyler Display Case

(Courtesy of Tyler) .................................................... 35 Figure 18. Two-minute Profile of the Controlled Environment Room

Dry Bulb and Relative Humidity Over 24 Hours – Hill Phoenix Display Case ................................................. 36

Figure 19. Two-minute Profile of Suction and Discharge Pressures Over 24 Hours – Hill Phoenix Display Case .................... 37

Figure 20. Two-minute Profile of Average Discharge and Return Air Temperatures Over 24 Hours – Hill Phoenix Display case ........................................................................ 37


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Figure 21. Two-minute Profile of Collected Condensate Over 24 Hours – Hill Phoenix Display Case ................................ 38

Figure 22. Breakdown of Condensate Collected Over 24 Hours – Hill Phoenix Display Case ............................................ 38

Figure 23. Two-minute Profile of Display Case Temperature and Relative Humidity Over 24 Hours – Hill Phoenix Display Case ........................................................................ 39

Figure 24. Hourly Profile of Total Cooling Load per Linear Foot of the Display Case Over 24 Hours – Hill Phoenix Display Case ........................................................................ 39

Figure 25. Cooling Load by Component Over 24 Hours – Hill Phoenix Display Case ................................................. 40

Figure 26. Percentage Breakdown of the Cooling Load Components Over 24 hours – Hill Phoenix Display Case .................... 40

Figure 27. Reduced Cooling Load, and Average Cooling Load Over 24 Hours and ¾ of Running Cycle – Hill Phoenix Display Case ........................................................................ 41

Figure 28. Two-minute Profile of Refrigerant Mass Flow Rate Over 24 Hours – Hill Phoenix Display Case ............................ 41

Figure 29. Two-minute Profile of Compressor Power and Refrigerant Mass Flow Rate Over 24 Hours – Hill Phoenix Display Case ................................................. 42

Figure 30. Hourly Profile of Evaporator Coil Temperature Difference (TD) Over 24 Hours – Hill Phoenix Display Case ........................................................................ 42

Figure 31. Hourly Profile of Evaporator Coil Superheat and Total System Sub-cooling Over 24 Hours – Hill Phoenix Display Case ............................................................. 43

Figure 32. Two-minute Profile of Case Lighting and Evaporator Fan Motor Power Over 24 Hours – Hill Phoenix Display Case .. 43

Figure 33. Average Total and End-use Power Over 24 Hours – Hill Phoenix Display Case ................................................. 44

Figure 34. Two-minute Profile of Product Temperature at Six Different Locations for Bottom Shelf Over 24 Hours – Hill Phoenix Display Case ............................................ 44

Figure 35. Two-minute Profile of Product Temperature at Six Different Locations for Second Shelf Over 24 Hours – Hill Phoenix Display Case ............................................ 45

Figure 36. Two-minute Profile of Product Temperature at Six Different Locations for Third Shelf Over 24 Hours – Hill Phoenix Display Case ................................................. 45

Figure 37. Two-minute Profile of Product Temperature at Six Different Locations for Fourth Shelf Over 24 Hours – Hill Phoenix Display Case ................................................. 46


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Figure 38. Two-minute Profile of Product Temperature at Six Different Locations for Top Shelf Over 24 Hours – Hill Phoenix Display Case ................................................. 46

Figure 39. Average Product Temperatures for Each Shelf Over 24 Hours – Hill Phoenix Display Case ................................ 47

Figure 40. Average, Coldest and Warmest Product Temperatures Over 24 Hours – Hill Phoenix Display Case .................... 47

Figure 41. Two-minute Profile of the Controlled Environment Room Dry Bulb and Relative Humidity Over 24 Hours – Hussmann Display Case ............................................. 49

Figure 42. Two-minute Profile of Suction and Discharge Pressures Over 24 Hours – Hussmann Display Case ...................... 49

Figure 43. Two-minute Profile of Average Discharge and Return Air Temperatures Over 24 Hours – Hussmann Display case .. 50

Figure 44. Two-minute Profile of Collected Condensate Over 24 Hours – Hussmann Display Case .................................. 50

Figure 45. Breakdown of Condensate Collected Over 24 Hours – Hussmann Display Case ............................................. 51

Figure 46. Two-minute Profile of Display Case Temperature and Relative Humidity Over 24 Hours – Hussmann Display Case ........................................................................ 51

Figure 47. Hourly Profile of Total Cooling Load per Linear Foot of the Display Case Over 24 Hours – Hussmann Display Case ........................................................................ 52

Figure 48. Cooling Load by Component Over 24 Hours – Hussmann Display Case ............................................................. 52

Figure 49. Percentage Breakdown of the Cooling Load Components Over 24 hours – Hussmann Display Case ...................... 53

Figure 50. Reduced Cooling Load, and Average Cooling Load Over 24 Hours and ¾ of Running Cycle – Hussmann Display Case ........................................................................ 53

Figure 51. Two-minute Profile of Refrigerant Mass Flow Rate Over 24 Hours – Hussmann Display Case ............................. 54

Figure 52. Two-minute Profile of Compressor Power and Refrigerant Mass Flow Rate Over 24 Hours – Hussmann Display Case ............................................................. 54

Figure 53. Hourly Profile of Evaporator Coil Temperature Difference (TD) Over 24 Hours – Hussmann Display Case ........................................................................ 55

Figure 54. Hourly Profile of Evaporator Coil Superheat and Total System Subcooling Over 24 Hours – Hussmann Display Case ........................................................................ 55

Figure 55. Two-minute Profile of Case Lighting and Evaporator Fan Motor Power Over 24 Hours – Hussmann Display Case .... 56


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Figure 56. Average Total and End-use Power Over 24 Hours – Hussmann Display Case ............................................. 56

Figure 57. Two-minute Profile of Product Temperature at Six Different Locations for Bottom Shelf Over 24 Hours – Hussmann Display Case ............................................. 57

Figure 58. Two-minute Profile of Product Temperature at Six Different Locations for Second Shelf Over 24 Hours – Hussmann Display Case ............................................. 57

Figure 59. Two-minute Profile of Product Temperature at Six Different Locations for Third Shelf Over 24 Hours – Hussmann Display Case ............................................. 58

Figure 60. Two-minute Profile of Product Temperature at Six Different Locations for Top Shelf Over 24 Hours – Hussmann Display Case ............................................. 58

Figure 61. Average Product Temperatures for Each Shelf Over 24 Hours – Hussmann Display Case .................................. 59

Figure 62. Average, Coldest and Warmest Product Temperatures Over 24 Hours – Hussmann Display Case ...................... 59

Figure 63. Two-minute Profile of the Controlled Environment Room Dry Bulb and Relative Humidity Over 24 Hours – Tyler Display Case ............................................................. 61

Figure 64. Two-minute Profile of Suction and Discharge Pressures Over 24 Hours – Tyler Display Case ............................. 61

Figure 65. Two-minute Profile of Average Discharge and Return Air Temperatures Over 24 Hours – Tyler Display case .......... 62

Figure 66. Individual and Average Discharge Air Temperature Over 24 Hours – Tyler Display case ..................................... 62

Figure 67. Two-minute Profile of Collected Condensate Over 24 Hours – Tyler Display Case ......................................... 63

Figure 68. Breakdown of Condensate Collected Over 24 Hours – Tyler Display Case ..................................................... 63

Figure 69. Two-minute Profile of Display Case Temperature and Relative Humidity Over 24 Hours – Tyler Display Case .... 64

Figure 70. Hourly Profile of Total Cooling Load per Linear Foot of the Display Case Over 24 Hours – Tyler Display Case ..... 64

Figure 71. Cooling Load by Component Over 24 Hours – Tyler Display Case ............................................................. 65

Figure 72. Percentage Breakdown of the Cooling Load Components Over 24 hours – Tyler Display Case .............................. 65

Figure 73. Reduced Cooling Load, and Average Cooling Load Over 24 Hours and ¾ of Running Cycle – Tyler Display Case ... 66

Figure 74. Two-minute Profile of Refrigerant Mass Flow Rate Over 24 Hours – Tyler Display Case ..................................... 66


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Figure 75. Two-minute Profiles of Compressor Power and Refrigerant Mass Flow Rate Over 24 Hours – Tyler Display Case ............................................................. 67

Figure 76. Hourly Profile of Evaporator Coil Temperature Difference (TD) Over 24 Hours – Tyler Display Case ....... 67

Figure 77. Hourly Profile of Evaporator Coil Superheat and Total System Subcooling Over 24 Hours – Tyler Display Case .. 68

Figure 78. Hourly Profile of Case Lighting and Evaporator Fan Motor Power Over 24 Hours – Tyler Display Case ........... 68

Figure 79. Average Total and End-use Power Over 24 Hours – Tyler Display Case ..................................................... 69

Figure 80. Two-minute Profile of Product Temperature at Six Different Locations for Bottom Shelf Over 24 Hours – Tyler Display Case ..................................................... 69

Figure 81. Two-minute Profile of Product Temperature at Six Different Locations for Second Shelf Over 24 Hours – Tyler Display Case ..................................................... 70

Figure 82. Two-minute Profile of Product Temperature at Six Different Locations for Third Shelf Over 24 Hours – Tyler Display Case ..................................................... 70

Figure 83. Two-minute Profile of Product Temperature at Six Different Locations for Fourth Shelf Over 24 Hours – Tyler Display Case ..................................................... 71

Figure 84. Two-minute Profile of Product Temperature at Six Different Locations for Top Shelf Over 24 Hours – Tyler Display Case ............................................................. 71

Figure 85. Average Product Temperatures for Each Shelf Over 24 Hours – Tyler Display Case ......................................... 72

Figure 86. Average, Coldest and Warmest Product Temperatures Over 24 Hours – Tyler Display Case ............................. 72

Figure 87. Comparison of Two-minute Profiles of the Controlled Environment Room Dry Bulb and Relative Humidity Over 24 Hours – All Three Test Scenarios ..................... 74

Figure 88. Comparison of Two-minute Profiles of Suction and Discharge Pressures Over 24 Hours – All Three Test Scenarios ................................................................. 74

Figure 89. Comparison of Two-minute Profiles of Average Discharge and Return Air Temperatures Over 24 Hours – All Three Test Scenarios .......................................... 75

Figure 90. Comparison of Two-minute Profiles of Average Discharge Air Temperature and Product Temperature Over 24 Hours – All Three Test Scenarios ..................... 76

Figure 91. Comparison of Coldest and Warmest Product Temperatures Over 24 Hours – All Three Test Scenarios . 77


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Figure 92. Comparison of Two-minute Profiles of Refrigeration Effect and Refrigerant Mass Flow Rate Over 24 Hours – All Three Test Scenarios ............................................. 77

Figure 93. Comparison of Two-minute Profiles of Compressor Power and Refrigerant Mass Flow Rate Over 24 Hours – All Three Test Scenarios ............................................. 78

Figure 94. Comparison of Two-minute Profiles of Mass of Collected Condensate Over 24 Hours – All Three Test Scenarios .... 79

Figure 95. Comparison of Total Cooling Load and Its Components Over 24 Hours – All Three Test Scenarios ..................... 80

Figure 96. Comparison of Test Data and Manufacturer’s Reported Cooling Load per Linear-feet of the Display Case – All Three Test Scenarios ................................................. 80

Figure 97. Comparison of Total and End-use Power Over 24 Hours – All Three Test Scenarios .......................................... 81

Figure 98. Comparison of Total Daily Defrost Periods and Refrigeration (compressor) Run Time Over 24 Hours – All Three Test Scenarios ............................................. 82

Figure 99. Comparison of Total and End-use Daily Energy Over 24 Hours – All Three Test Scenarios ................................. 83

Figure 100. Comparison of Total Cooling Load and Power per Refrigerated Volume – All Three Test Scenarios ............. 83

TABLES Table 1. Lineup Length, Suction Temperature Group, and

Cooling Load by Type of Open Vertical Multi-deck Refrigerated Display Case in a Typical Supermarket ......... 5

Table 2. Specification Summary of Tested Display Cases ............. 10 Table 3. Specifications of Sensors Used .................................... 13 Table 4. Comparative Summary of Test Data and Manufacturer’s

Published Data – Hill Phoenix Display Case .................... 48 Table 5. Comparative Summary of Test Data and Manufacturer’s

Published Data – Hussmann Display Case ..................... 60 Table 6. Comparative Summary of Test Data and Manufacturer’s

Published Data – Tyler Display Case ............................. 73 Table 7. Summary of Key System Parameters and Measured

Condensate for All Three Test Scenarios ....................... 84 Table 8. Summary of Key Refrigeration Parameters and Cooling

Load for All Three Test Scenarios ................................. 84 Table 9. Summary of Power Demand and Daily Energy Usage for

All Three Test Scenarios ............................................. 85


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EQUATIONS Equation 1. Refrigeration Effect .................................................... 17 Equation 2. Total Refrigeration Load of the Display Case (in btu/hr) .. 17 Equation 3. Total Refrigeration Load of the Display Case (in cooling

tons) ....................................................................... 17 Equation 4. Volumetric Flow Rate of Air Into the Display Case .......... 18 Equation 5. Mass Flow Rate of Air ................................................. 18 Equation 6. Mass of Condensate Collected From Air During Defrost

Period ...................................................................... 18 Equation 7. Mass of Melted Frost During Defrost Period ................... 19 Equation 8. Sensible Load of Refrigeration ..................................... 19 Equation 9. Latent Load of Refrigeration ........................................ 20 Equation 10. Cooling Load During the Last Three-Quarters of the

Refrigeration Run Cycle .............................................. 20 Equation 11. Reduction Factor for Refrigeration Run Cycle ............ 20 Equation 12. Cooling Load for one Refrigeration Run Cycle ............ 21 Equation 13. Temperature Differential (T) Across the Evaporator

Coil ......................................................................... 21 Equation 14. Temperature Difference (TD) Across the Evaporator

Coil ......................................................................... 21 Equation 15. Evaporator Coil Superheat ..................................... 21 Equation 16. Evaporator Coil Moisture Removal Rate ................... 22 Equation 17. Evaporator Coil Log-Mean Temperature Difference

(LMTD) .................................................................... 22 Equation 18. Evaporator Coil Effective Overall Heat Transfer

Coefficient (UA) ........................................................ 22 Equation 19. Total Refrigeration Power Usage, Excluding

Condenser ................................................................ 22 Equation 20. Energy Usage by the Evaporator Fan Motors ............. 23 Equation 21. Energy Usage by the Secondary Fan Motors ............. 23 Equation 22. Energy Usage by the Light Fixtures in the Display

Case ........................................................................ 23 Equation 23. Energy Usage by the Compressor ........................... 24 Equation 24. Total Refrigeration Energy Usage, Excluding

Condenser ................................................................ 24 Equation 25. Overall Heat Transfer Coefficient for the Display

Case Walls ............................................................... 25


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Equation 26. Transmission or Conduction Load of the Display Case ........................................................................ 25

Equation 27. Radiation Load of the Display Case ......................... 27 Equation 28. Display Case Load due to Evaporator Fan Motors ...... 28 Equation 29. Display Case Load due to Lighting ........................... 28 Equation 30. Infiltration Load of the Display Case ........................ 28 Equation 31. Volumetric Flow Rate of Infiltrated Air From Room

Into the Display Case ................................................. 29 Equation 32. Sensible Portion of the Infiltration Load of the

Display Case ............................................................. 29 Equation 33. Latent Portion of the Infiltration Load of the Display

Case ........................................................................ 29


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CONTENTS EXECUTIVE SUMMARY _______________________________________________ 1

INTRODUCTION ____________________________________________________ 3 Background ........................................................................... 3 Goals and Objectives .............................................................. 7

TECHNICAL APPROACH _____________________________________________ 8

TEST FACILITY _____________________________________________________ 9

TEST DESIGN AND INSTRUMENTATION ___________________________________ 10 Test Design ......................................................................... 10 Instrumentation ................................................................... 12

DATA ACQUISITION, DATA COLLECTION AND SCREENING PROCEDURE ________ 15 Data Acquisition ................................................................... 15 Data Collection and Screening Procedure ................................. 16

DATA ANALYSIS __________________________________________________ 16 Refrigeration Cycle Analysis ................................................... 16

Refrigeration Effect .......................................................... 17 Refrigeration Load ........................................................... 17 Airflow Rate .................................................................... 17 Mass of Condensate ......................................................... 18 Sensible and Latent Loads ................................................ 19 Cooling Load Based on One Running Cycle .......................... 20 Evaporator Coil Characteristic Performance ......................... 21 Total System Power and Energy ........................................ 22

Display Case Heat Transfer Analysis ........................................ 24 Transmission (or Conduction) Load .................................... 24 Radiation Load ................................................................ 26 Internal Load .................................................................. 27 Infiltration Load .............................................................. 28

DESCRIPTION OF DISPLAY CASES _____________________________________ 30 Hill Phoenix Display Case – O5DM ........................................... 30 Hussmann Display Case – M5X-GEP ........................................ 32 Tyler Display Case – N6DHPACLA ........................................... 34

RESULTS ________________________________________________________ 36 Hill Phoenix Display Case (O5DM) ........................................... 36


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Hussmann Display Case (M5X-GEP) ........................................ 48 Tyler Display Case (N6DHPACLA) ............................................ 60

COMPARISON OF RESULTS __________________________________________ 73

CONCLUSIONS AND RECOMMENDATIONS ______________________________ 86

REFERENCES _____________________________________________________ 87


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EXECUTIVE SUMMARY This Emerging Technology project was conducted to quantify and compare the key performance attributes of a new generation of high efficiency medium-temperature open vertical refrigerated display cases. The objective of this laboratory assessment was to determine the power and energy implications of using the latest commercially available energy efficient medium-temperature display cases. The benefits of using these high efficiency display cases was evaluated by measuring key performance parameters such as cooling load, product temperatures, and compressor power and energy requirements.

This project evaluated three high efficiency medium-temperature open vertical refrigerated display cases from three leading U.S. display case manufacturers, namely Hill Phoenix, Hussmann, and Tyler. The primary selection criterion was the classification, the similarities in physical characteristics, and the application of these cases. All three acquired display cases were standard high efficiency models without any extra options or features. The tested display case manufacturers and their corresponding product specifications are detailed in this report.

A comprehensive monitoring plan was developed to ensure all critical data points were captured. The monitoring involved measuring cooling load, product temperatures, and power and energy usage of end-use components, to name a few. The monitoring also involved measuring and tracking control variables like discharge air temperature, saturated evaporating temperature, and saturated condensing temperature.

After data was screened and sanitized, data analysis took place. Data analysis included refrigeration cycle and heat transfer analysis. After the collected data was analyzed, the findings were shared and discussed with the manufacturer representatives. This was an important step in the project to ensure the findings were in line with the manufacturer’s expectations.

The results of this study indicated that the total cooling load of the open vertical refrigerated display case with the lowest vertical distance between the discharge and return air grille was 22% lower than the other two display cases. Because the infiltration load contributed to more than 80% of the total cooling load of these cases, the variations in total cooling load was attributed to variations in infiltration load. In fact, the infiltration load of the Hussmann case was 26% lower than the Hill Phoenix case and 12% lower than the Tyler case. Due to a larger surface area of the case walls, the Hill Phoenix case had the highest conduction load (637 Btu/hr) when compared to the Hussmann (551 Btu/hr) and Tyler case (496 Btu/hr). The radiation load, however, remained fairly unchanged around 1,000 Btu/hr for all three display cases. The internal load, which was comprised of heat generated by the case lighting system and evaporator fan motors, was higher for the Hussmann case (730 Btu/hr) when compared to the Hill Phoenix (592 Btu/hr) and Tyler case (476 Btu/hr). This was attributed mainly to an increase in evaporator fan motor power of the Hussmann case prior to initiation of defrosts.

Due to lower cooling load requirements of the Hussmann case, the refrigeration compressor required less power to provide or satisfy the cooling load of this case. The compressor power demand for the Hussmann case was 14% lower than the Hill Phoenix case and 9% lower than the Tyler case. Since the compressor was turned off during defrost periods, the compressor run time was a function of defrost frequency and duration. The compressor daily run time was about 22 hours for both the Hussmann and Tyler display case test scenarios, and about 21 hours for the Hill Phoenix display case test scenario. As expected, the compressor daily energy usage followed a similar pattern as the power demand because the



run times were not significantly different for all three tested display cases. The compressor consumed about 11% to 12% less energy per day during Hussmann display case test scenario when compared to the other two test scenarios or display cases.

Comparison of cooling load and power demand per refrigerated volume of display cases showed that the Hill Phoenix case had the highest cooling load requirement per refrigerated volume (175 Btu/hr/ft3) followed by the Hussmann (150 Btu/hr/ft3) and Tyler (147 Btu/hr/ft3) cases. Similar observations were made regarding the total power demand per refrigerated volume. In other words, per refrigerated volume of the case, the Tyler display case had the lowest cooling load and power demand requirements whereas the Hill Phoenix case had the highest cooling load and power demand requirements.

Finally, comparing the coldest and warmest product temperatures revealed that the coldest product temperature for all three tested display cases was between 27oF and 34oF. More importantly, the warmest product temperature for both Hill Phoenix and Tyler cases was above the Food and Drug Administration’s food code requirement of 41oF. This difference was more pronounced for the Tyler case, with a 7oF difference, than for the Hill Phoenix case, with less than 1oF difference. Nonetheless, the warmest product temperature for the Hussmann case was about 40oF, which was 1oF lower than the Food and Drug Administration’s food code requirement.

In summary, the Hussmann display case had the lowest cooling load requirement, and specifically infiltration load. This in turn, resulted in lower power demand and energy usage. More importantly, lower power and energy usage were achieved while maintaining the warmest product temperatures below 41oF.

Based on these findings, it was recommended to select open vertical refrigerated display cases with following characteristics, while maintaining the warmest product temperature equal to or below 41oF:

Lowest temperature difference between the discharge and return air (below 10oF)

Lowest vertical distance between the discharge and return air grille

Least amount of daily collected condensate or defrost water (below 9.5 lb/ft/day)

Lowest infiltration load per refrigerated volume (below 120 Btu/hr/ft3)

Lowest total cooling load per refrigerated volume (below 145 Btu/hr/ft3)

Lowest evaporator fan motor power (below 20 watts/fan motor)

Lowest display case lighting power (below 55 watts/canopy row)



INTRODUCTION This technology assessment investigated the demand and energy usage of a new generation of high efficiency medium-temperature (MT) open vertical refrigerated display cases (OVRDCs). Three new generation high efficiency OVRDCs were evaluated from three of the leading U.S. manufacturers, namely Hill Phoenix, Hussmann, and Tyler. The evaluation involved measuring key performance parameters such as cooling load, product temperatures, and compressor power and energy requirements.

Medium-temperature OVRDCs have a large presence in supermarkets and account for more than 50% of total display case lineups. Since these display cases are commonly used to merchandise meat, dairy, deli, produce and fish, their operation is especially critical because the Food and Drug Administration (FDA) strictly regulates the temperature of these products. These cases contribute to roughly 60% of refrigeration energy use in a typical supermarket.

BACKGROUND Supermarkets and grocery stores represent one of the largest electric energy-intensive building groups in the commercial sector, at 43 to 70 kWh/ft2 per year [Ref. 1]. A typical 50,000 ft2 supermarket, which is classified as large supermarket, consumes somewhere between 2 to 3 million kWh per year [Ref. 2]. About 50% of this energy use, however, is for the refrigeration of food display cases and storage coolers [Ref. 1]. Based on commercial end-use survey data, it is estimated that there are roughly 6,900 and 2,800 supermarkets with annual energy consumption of greater than 1.6 million kWh in the State of California and Southern California Edison’s (SCE’s) service territory, respectively [Ref. 3].

Display cases are widely used in supermarkets and grocery stores for merchandising of perishable food products. Depending upon the type of product stored, hence temperature requirements, display cases can be categorized as either medium- or low-temperature. To maintain proper and desired product temperatures, display cases rely heavily on the temperature of air discharged into the case or the discharge air temperature (DAT). For example, MT display cases are used to merchandise meat, deli, dairy, produce and beverages. The DAT of these types of display cases can range from +24oF to +38oF [Ref. 1]. Low-temperature (LT) display cases, on the other hand, are used to merchandise frozen food and ice cream. The DAT for LT display cases can range from -24oF to -5oF [Ref. 1]. Figure 1 illustrates the distribution of display cases by type in a typical supermarket. As shown, about half of the total refrigerated display cases in a supermarket are MT open vertical multi-deck [Refs. 1, 2].



Low TempReach-ins

33%

Medium-Temp Reach-ins

1%Medium-TempMulti-Deck

Open Vertical50%

Medium-Temp Single-DeckOpen Cases

3%

Medium-TempIsland Cases

11%

Medium-TempService Cases

4%

FIGURE 1. PERCENTAGE BREAKDOWN OF DISPLAY CASES BY TYPE IN A TYPICAL SUPERMARKET [REF 2]

Table 1 shows the lineup length (in linear-feet), the corresponding suction temperature group (in oF), and the refrigeration or cooling load (in Btu/hr) for each type of OVRDC. Specifically, Table 1 shows that there are about 330 linear-feet of OVRDCs in a typical 50,000 ft2 supermarket totaling over 490,000 Btu/hr, or 41 tons of refrigeration load. It is common practice to select refrigeration compressors using a 15% over-sizing factor. Therefore, the required compressor capacity for the MT refrigeration system will yield 565,041 Btu/hr, or 47 tons of refrigeration load. To satisfy this cooling load, the equivalent-full-load-hours (EFLH) of operation of refrigeration compressors serving OVRDCs is 6,398 hours per year, which was established based on the electric billing data for a typical supermarket. Further, using compressor manufacturers catalog data and design saturated condensing temperature (SCT) of 90oF for refrigerant R-404A, the energy-efficiency ratio (EER) of these compressors is estimated to be around 12.5 Btu/hr/watt. Accordingly, it can be estimated that the refrigeration compressors of a typical supermarket require about 46 kW and 290,000 kWh per year to remove 565,041 Btu/hr or 47 tons of refrigeration load.

Subsequently, the power demand and energy usage of MT refrigeration compressors of 2,800 large supermarkets in the SCE’s service territory can be estimated to be about 128 MW and 812 GWh, respectively. Similarly, the power demand and energy usage of MT refrigeration compressors of 6,900 large supermarkets in the State of California can be estimated to be about 317 MW and 2,001 GWh, respectively.



TABLE 1. LINEUP LENGTH, SUCTION TEMPERATURE GROUP, AND COOLING LOAD BY TYPE OF OPEN VERTICAL MULTI-DECK REFRIGERATED DISPLAY CASE IN A TYPICAL SUPERMARKET

MEDIUM-TEMPERATURE OPEN VERTICAL

MULTI-DECK DISPLAY CASE TYPES

LINEUP LENGTH (LINEAR-FEET)

SUCTION TEMPERATURE GROUP (OF)

REFRIGERATION OR COOLING LOAD

(BTU/HR)

Fresh Meat 59 +15 78,175

Dairy 62 +20 92,690

Deli 56 +15 83,720

Beverage 55 +20 82,225

Produce 102 +20 154,530

Total 334 491,340

The schematics of a typical OVRDC (side view) and the air circulation pattern for these display cases is shown in Figure 2. As shown, cold air is provided through an inlet jet called the discharge air grille (DAG) located at the top front of the case, and through a group of slots located on the back panel of the case. The air is re-circulated to the evaporator for cooling through an outlet located at the bottom front of the case called the return air grille (RAG). This top-down flow of cold air creates an invisible barrier between the refrigerated space and the warm and moist adjacent space, and is called the air curtain. However, the mixing between the cold and warm air cannot be avoided when part of the cold air spills over the display case and is replaced by warm air. The continuous flow of warm air into the air curtain and its subsequent mixing with cold air is called entrainment. A portion of the entrained air spills over after some mixing with the cold air, and the rest is infiltrated into the RAG. The amount of warm and moist air that moves into the thermodynamic cooling cycle of the display case through the RAG is called the infiltration rate, and it is responsible for the infiltration load of an OVRDC. The infiltration load accounts for most of the cooling load of an OVRDC and thereby power consumption.



FIGURE 2. SCHEMATICS OF A TYPICAL OPEN VERTICAL REFRIGERATED DISPLAY CASE AND AIR CIRCULATION PATTERN (SIDE VIEW)

The total cooling load of an OVRDC is comprised of four distinct sources: (1) heat conduction through the case panels, (2) thermal radiation from the adjacent space to the display case interior, (3) internal thermal loads such as case lighting, evaporator fan motors, and period defrosts, and (4) infiltration of warm and moist air from the adjacent space into the display case through the RAG. As shown in Figure 3, infiltration through the air curtain plays a significant role in the cooling load of OVRDCs and constitutes roughly 80% of the total cooling load [Refs. 1, 2]. The remaining 20% of the total cooling load is comprised of conduction, radiation, and thermal loads due to case lighting and evaporator fan motors [Refs. 1, 2].



Infiltration81%

Radiation8%

Conduction3%

Case Lighting

6%

EvaporatorFans

3%

FIGURE 3. REFRIGERATION LOAD FOR FOR TYPICAL MEDIUM-TEMPERATURE OPEN VERTICAL REFRIGERATED DISPLAY CASE AT 75OF DRY BULB AND 55% RELATIVE HUMIDITY [REF 1]

GOALS AND OBJECTIVES This laboratory assessment project determined the power and energy implications of using the latest commercially available energy efficient MT OVRDCs. The benefits of using energy efficient display cases were evaluated by measuring key performance components such as cooling load, product temperature, and compressor power and energy requirements.



TECHNICAL APPROACH The following lists the necessary steps taken from the start to the conclusion of the project. A brief discussion on each of these milestones is presented in this section.

1. Select three MT OVRDCs

2. Develop monitoring plan

3. Install sensors and data acquisition equipment

4. Collect monitoring data

5. Reduce and screen data

6. Develop engineering analysis tool and analyze data

7. Share findings with all three case manufacturers

8. Prepare and finalize report

The high efficiency MT OVRDCs from three leading display case manufacturers were selected. All three acquired display cases were standard high efficiency models without any added options or features. The primary selection criterion was the classification, the similarities in physical characteristics, and the application of these cases. Evaluating three different cases will enhance understanding about the variation in design and performance of these cases.

A comprehensive monitoring plan was developed to ensure all critical data points were captured. The monitoring involved measuring key performance components such as cooling load, product temperature, and compressor power and energy requirements. The monitoring also involved measuring and tracking control variables such as discharge air temperature (DAT), saturated evaporating temperature (SET) and saturated condensing temperature (SCT).

After data was screened and sanitized, data analysis took place. Data analysis included comparing cooling load, power and energy consumption of the high efficiency MT OVRDCs. Data analysis also included comparing power and energy as a function of total refrigerated volume.

After the collected data was analyzed, the findings were shared and discussed with the manufacturer’s representatives. This was an important step in the project to ensure the findings were in line with the manufacturer’s expectations.



TEST FACILITY This laboratory test project was conducted at Southern California Edison’s (SCE’s) Technology Test Centers (TTC). The TTC is a 7,500 square-feet testing facility located in Irwindale, California. The TTC is comprised of two main centers:

1. Southern California Lighting Technology Center (SCLTC) – focusing on lighting technologies and applications

2. Refrigeration and Thermal Test Center (RTTC) – focusing on refrigeration and HVAC related technologies and applications

The display cases were tested in the controlled environment room of the RTTC. This room is an isolated thermal zone served by independent cooling, heating and humidification systems. This allows simulation of various indoor conditions of a supermarket. The sensible cooling load representing people and other heat gain sources is provided by a constant volume direct expansion system reclaiming the waste refrigeration heat via a six-row coil. Auxiliary electric heaters located downstream of the heat reclaim coils provide additional heating, when required. While the air is conditioned to a desired thermostatic set point, an advanced ultrasonic humidification unit introduces precise amounts of moisture to the air surrounding the display cases, representing the latent load due to outside air and people. Figure 4 shows a schematic diagram of the air conditioning and heating system of the RTTC’s controlled environment.

FIGURE 4. SCHEMATIC DIAGRAM OF THE AIR CONDITIONING AND HEATING SYSTEM OF THE RTTC’S CONTROLLED ENVIRONMENT ROOM

There are three laminar diffusers in the room, each supplying air at approximately 370 cubic feet per minute (cfm). The intensity of ambient lighting in the controlled environment room, as measured from the center of the test fixture opening at a distance of 1 foot from the air curtain, is 115 foot-candles. This meets American Society of Heating, Refrigeration and Air-Conditioning Engineers (ASHRAE) Standard 72-05, which requires the lighting intensity not be less than 75 foot-candles at this location.



TEST DESIGN AND INSTRUMENTATION Thermal testing and analysis was carried out to quantify the performance of three standard high efficiency MT OVRDCs. In this context, display case performance refers to the total refrigeration or cooling load, cooling load components, power and energy usage, and product temperatures. Thermal testing followed test procedures and guidelines specified in ASHRAE Standard 72-05. Based on manufacturers’ data, a specific summary of three tested display cases is provided in Table 2.

TABLE 2. SPECIFICATION SUMMARY OF TESTED DISPLAY CASES

DISPLAY CASE TYPE

MAKE MODEL APPLICATION DISCHARGE AIR TEMP.

(OF)

CAPACITY (BTU/HR/FT)*

LENGTH (FOOT)

Open 5-Deck Front Loading

Hill Phoenix O5DM Deli (MT) 30.0 1,570 8


Hussmann M5X-GEP Meat and Deli

(MT) 30.0 1,370 8


Tyler N6DHPACLA Dairy (MT) 34.5 1,059 8

* Btu/hr/ft listed conventional ratings.

TEST DESIGN All tests were performed under steady-state conditions following ASHRAE Standard 72-05. The refrigeration system was charged with a hydrofluorocarbon refrigerant (R-404A). The refrigeration system controller maintained a fixed saturated condensing temperature (SCT) of 95oF + 0.5oF for all tests. To comply with manufacturers’ specifications for performance evaluations, the average discharge air temperature (DAT), which was the critical control point, was maintained at their specified temperatures (see Table 2).

The controlled environment chamber was maintained at a constant dry bulb (DB) temperature of 75.2oF + 2oF and wet bulb of 64.4oF + 2oF, corresponding to 55% relative humidity (RH), throughout the entire 24-hour test period. The intensity of ambient lighting in the controlled environment room was 115 foot-candles and was in compliance with the ASHRAE standard, which requires a minimum of 74.4 foot-candles. The foot-candle measurement was taken at a distance of one foot from the air curtain. The entering liquid refrigerant temperature and pressure, measured at 6.1 feet of pipe length from the display case, were maintained at 80oF and 214 psig (corresponding to an SCT of ~94oF). These parameters meet the ASHRAE standard, which requires the entering liquid temperature be 80.6oF + 5oF and SCT be maintained between 89.6oF and 120.2oF.

The display case was mounted on a special platform to allow installation of a customized condensate pipe/valve arrangement. The piping and valve assembly transferred condensate from the fixture into the container placed on the digital scale. Figure 5 shows the fixture with this custom drainage assembly.



FIGURE 5. CUSTOM RAISED FRAME ASSEMBLY AND SPECIAL DRAIN PIPING/VALVE ARRANGEMENT

ASHRAE 72-05 also requires food product zones be filled with test packages and dummy products to simulate the presence of food product in the display cases (Figure 6). According to ASHRAE standard, food products are comprised of 80% to 90% water, fibrous materials, and salt. Therefore, plastic containers completely filled with a sponge material soaked in a 50% + 2% by volume solution of propylene glycol and distilled water were used to simulate the product during the tests. The spaces in the test display case where temperature measurement was not required were stocked with dummy products to stabilize the temperature in the case and account for transient heat transfer effects.

FIGURE 6. SIMULATED AND DUMMY PRODUCTS USED IN THE DISPLAY CASE

For each display shelf, six product simulators were used to monitor the product temperatures (Figure 7). Two product simulators were located at the left end, the right end and the center. At each left, right, and center location, one product simulator was placed on the shelf surface at the front of the shelf and one at the rear edge of the shelf.

product simulator

dummy products



MAIN CHANNELS DESCRIPTION

E FRONT PRODUCT TEMPERATURES

F BACK PRODUCT TEMPERATURES

FIGURE 7. LOCATION OF PRODUCT SIMULATORS INSIDE THE DISPLAY CASE

The test was designed with a strong emphasis on proper equipment set up, instrumentation, and data acquisition of the test scenarios. Results obtained from all tests addressed the following key parameters:

Compressor power and energy, (kW, kWh)

Total system power and energy (less condenser), (kW, kWh)

Evaporator fan motors power and energy, (kW, kWh)

Display case lighting power and energy, (kW, kWh)

Refrigeration energy, (Btu)

Case total cooling load, (Btu/hr)

Condensate quantity, (lbs/hr, lbs)

Product temperatures, (oF)

INSTRUMENTATION All temperature and pressure instruments were calibrated before the test. Careful attention was paid to the design of the monitoring system, with the objective of minimizing instrument error and maintaining a high level of repeatability and accuracy in the data. The monitoring plan was developed based on these guidelines:

Use of sensors with the highest accuracy available

Minimization of sensor drift errors by use of redundant sensors

Use of calibration standard instruments of the highest accuracy

Elimination of interference from power conductors and high frequency signals by double-shielding sensor leads

The instrumentation system includes these items:

Special grade type-T thermocouples accurate to + 0.1oC

Precision 100 platinum resistance temperature device (RTD) inputs accurate to 0.01C

Analog inputs from pressure transducers



Dew point sensors

Flow meters

CT-transducers

A USB communication link was used to send one data report including instantaneous values of all data points every 10 seconds. Table 3 provides the specifications of the various sensors used in the RTTC’s refrigeration system for this test. Figure 8 shows the location of sensors within the test fixtures.

The RTTC data acquisition system was set up to scan and log 99 data channels in 10-second intervals. Collected data was screened closely to ensure the key control parameters were within acceptable ranges. In the event that any of the control parameters fell outside acceptable limits, the problem was flagged and a series of diagnostic investigations were carried out. Corrections were then made and tests were repeated as necessary. After the data passed the initial screening process, it was imported to RTTC’s customized refrigeration analysis model where detailed calculations were performed. The collected data points from the 10-second intervals were averaged into 2-minute and hourly values, where necessary, and used for a secondary screening of the results.

TABLE 3. SPECIFICATIONS OF SENSORS USED

SENSOR TYPE MAKE/MODEL ACCURACY [NIST TRACEABLE]

Humidity Vaisala HMP247 + (0.5 + 2.5% of the reading) %RH

Dew Point EDGETECH Model 2000 Dew Prime DF Dew Point Hygrometer – S2 Sensor

+ 0.2oC (+ 0.36oF)

Refrigerant Mass Flow Micro Motion Model DS065S + 0.2%

Power Ohio Semitronics Model PC5-062BX680

+ 0.5% F.S. (0.04 kW)

Power Ohio Semitronics Model P-143B + 1.0% F.S. (0.08 kW)

Pressure Setra Transducers Model C207-100 & 500 PSIG Pressure Ranges

+ 0.13%

Pressure Danfoss Transducers Model AKS32 0-500 PSIG

+ 0.2% F.S.

Temperature (RTD) Hy-Cal Engineering Model RTS-37-A-100

+ 0.01oC

Temperature (TC) Kaye Instruments T/W 50 through 80; Melt # 8032

+ 0.1oC

Scale HP-30K + 0.1 gram



MAIN CHANNELS DESCRIPTION

A DISCHARGE AIR

B RETURN AIR

C AIR ENTERING EVAPORATOR [After Fans]

D AIR LEAVING EVAPORATOR

E FRONT PRODUCT TEMPERATURES

F BACK PRODUCT TEMPERATURES

G REFRIGERANT AT COIL EXIT

H AIR INSIDE CASE CAVITY

I REFRIGERANT AT EXPANSION VALVE EXIT

J REFRIGERANT AT EXPANSION VALVE INLET

K CASE LIGHT - POWER

L CASE FAN MOTORS - POWER

SUBSCRIPTS (numerical)

1 TEMPERATURE

2 RELATIVE HUMIDITY

3 PRESSURE

4 DEWPOINT TEMPERATURE

SUBSCRIPTS (roman)

i LEFT

j MIDDLE

k RIGHT

ik BETWEEN RIGHT AND MIDDLE

jk BETWEEN LEFT AND MIDDLE

ijk COMBINATION OF RIGHT, MIDDLE, AND LEFT

FIGURE 8. LOCATION OF SENSORS FOR OPEN VERTICAL MULTI-DECK DISPLAY CASES



DATA ACQUISITION, DATA COLLECTION AND SCREENING PROCEDURE

DATA ACQUISITION The National Instruments’ SCXI data acquisition system was used to log the test data. The data acquisition system was set up to process 99 data channels in 10-second intervals. The data acquisition system was calibrated at the factory, and is traceable to the National Institute of Standards and Technology’s (NIST) standards. As part of the RTTC’s quality control protocol, the data acquisition system for the project was designed to be completely independent of the supervisory control computer. This approach was taken to ensure that the data collection was not compromised by the control sequence’s priority over data acquisition.

The data acquisition system sampled the scanned data every 10 seconds. The 10-second data was then saved to a file, which was closed at the end of each 24-hour period. The initial data was reviewed on site at the RTTC to ensure that the key control parameters were within acceptable ranges. In the event that any of the control parameters fell outside acceptable limits, the problem was flagged. In these cases, test runs were repeated until the problem was corrected. After the data passed the initial screening process, it was downloaded for further screening and processing.

The weight of condensate during each test scenario was measured using a high precision digital scale with + 0.1 gram accuracy (Figure 9). The data acquisition system received the exact condensate weight measurements from the digital scale every 10 seconds. In this way it was possible to closely monitor and distinguish between the moisture removal from the air during the refrigeration cycle and defrost periods. At the end of each test period, the condensate data was also aggregated into 2-minute, hourly and daily values.

FIGURE 9. HIGH PRECISION DIGITAL SCALE USED TO MEASURE THE WEIGHT OF CONDENSATE COLLECTED



DATA COLLECTION AND SCREENING PROCEDURE The RTTC’s sophisticated data acquisition system scanned 99 data channels 100 times and logged their averages in 10-second intervals. Every 24 hours during the test, the data was checked for consistency and accuracy. Consistently, the key operating parameters were also checked and deemed to be within acceptable limits before the next run was started.

The data was then downloaded and detailed calculations were performed. The collected data points from the 10-second intervals were averaged into 2-minute and hourly values, where necessary, and used for further screening of the results. The advantage of using hourly averages is that the data trends can still be displayed with an acceptable resolution while enabling the engineering model to generate relevant calculated hourly results (e.g., cooling load). After the hourly data was developed, it was imported to RTTC’s customized refrigeration analysis tool.

After the data was compiled into 2-minute and hourly averages within the engineering model, tabular and graphical representations of various correlations and calculated parameters were produced. Several graphs were created to initially screen the calculated results. All critical raw data was screened and validated at the end of each 24-hour test, prior to importing it to RTTC’s engineering model. After careful examination and upon validation of the initial screening plots, the informational plots were produced. This set provided relationships between calculated quantities. In cases where data flaws were detected, a series of diagnostic investigations were conducted, and through this process, corrections were made, and tests were repeated when necessary.

DATA ANALYSIS The data analysis included refrigeration cycle and heat transfer analysis. Refrigeration cycle analysis provided key refrigeration parameters such as refrigeration effect and cooling load. Heat transfer analysis quantified incoming heat from the surrounding area into the display case.

REFRIGERATION CYCLE ANALYSIS Using refrigeration data, a series of calculations were performed to obtain the key refrigeration parameters. Next, the data was downloaded from the data logger and the data of interest was extracted, followed by preliminary reductions and calculations. These calculations included averaging of temperature, pressure, refrigerant mass flow, and condensate weight.

The total cooling load of the display case can be determined based on the refrigeration effect and mass flow rate of refrigerant. Determination of refrigeration effect and other quantities, such as heat of compression and sub-cooling quantities depend on the refrigerant enthalpies at specific locations within the refrigerant lines. Enthalpies can be obtained either from the refrigerant manufacturer’s data at various temperatures and pressures, or calculated with respect to specific heat capacities and temperatures. In this analysis, the refrigerant enthalpies were obtained using XPropsTM refrigerant property program, version 1.5. XPropsTM and was also used to determine the saturated refrigerant temperatures based on collected temperature and pressure data.



REFRIGERATION EFFECT The refrigeration effect is the quantity of heat that each unit of mass of refrigerant (in this case pound of refrigerant) absorbs to cool the refrigerated space. It simply represents the capacity of the evaporator per pound of refrigerant. This quantity was derived by subtracting the refrigerant enthalpy at the evaporator inlet (before the expansion valve) from the slightly superheated refrigerant enthalpy at the outlet of the evaporator (Equation 1).

EQUATION 1. REFRIGERATION EFFECT

RE = hevap-out – hevap-in

where,

RE = Refrigeration effect of the refrigerant in the evaporator, (Btu/lb)

hevap-out = Superheated refrigerant enthalpy at the evaporator exit, (Btu/lb)

hevap-in = Sub-cooled liquid refrigerant enthalpy at expansion valve inlet, (Btu/lb)

REFRIGERATION LOAD The refrigeration load of the case is the rate of cooling or heat removal (in BTU) that takes place at the evaporator of the display case per hour (Equation 6). This quantity is obtained by multiplying the refrigeration effect by refrigerant mass flow rate, which is extracted from the data acquisition system. The total case load for the display case was determined by using Equation 2.

EQUATION 2. TOTAL REFRIGERATION LOAD OF THE DISPLAY CASE (IN BTU/HR)

kREmQ refrefcase

where,

refcaseQ

= Total refrigeration load of the case, sensible and latent, (Btu/hr)

refm

= Mass flow rate of refrigerant, (lb/min)

k = Conversion factor, (60 min/hr)

To determine the refrigeration load of the case in tons, it can be divided by 12,000, a conversion factor for Btu/hr to tons (Equation 3).

EQUATION 3. TOTAL REFRIGERATION LOAD OF THE DISPLAY CASE (IN COOLING TONS)

12,000Q refcase

)(tonsrefcaseQ

where,

)(tonsrefcaseQ

= Refrigeration load, (tons)

AIRFLOW RATE The psychrometric analysis relies heavily on the mass flow rate of the air within the thermodynamic boundary of the refrigerated fixture. The volume flow rate of air



circulated throughout the fixture is a required parameter for conducting psychrometric calculations. This parameter was obtained using an approximation approach. This approximation relied on the discharge air velocity, free area available at the discharge air grille, and perforations in the back panel of the display case (Equation 4).

EQUATION 4. VOLUMETRIC FLOW RATE OF AIR INTO THE DISPLAY CASE

avgDAGpanelbackcase DAVAAcfm where,

casecfm = Volumetric flow rate of air into the display case, (ft3/min)

panelbackA = Total area of openings in the back panel, (ft2)

DAGA = Total free area available through discharge air grille, (ft2)

avgDAV = Average discharge air velocity through discharge air grille, (ft/min)

After the volumetric flow rate of air into the display case was determined, the mass flow rate of air was obtained (Equation 5).

EQUATION 5. MASS FLOW RATE OF AIR

kcfmm inaircaseair

where,

airm

= Mass flow rate of air, (lb/hr)

inair = Density of air at the inlet of the evaporator coil, (lb/ft3)


MASS OF CONDENSATE Mass of condensate can be comprised of the following constituents:

1. Mass of water vapor condensed from air during the defrost period

2. Mass of water vapor condensed from air during the refrigeration period

3. Mass of melted frost during defrost

The different components of condensate mass were obtained using the following equations. The total mass and the portion of condensate collected during refrigeration were obtained directly from scale readings. Equation 6 used psychrometric data to differentiate the defrost portion from the rest of the condensate mass.

EQUATION 6. MASS OF CONDENSATE COLLECTED FROM AIR DURING DEFROST PERIOD

-air in air out air defrostcond def

refrig

m tm

t

where,



defcondm

= Mass of water vapor condensed from air during defrost period, (lb/hr)

inair = Absolute humidity of air at the evaporator inlet, (lbw/lba)

outair = Absolute humidity of air at the evaporator outlet, (lbw/lba)

defrostt = Defrost period, (hours)

refrigt = Refrigeration period, (hours)

Next, the mass of melted frost was determined. This quantity was determined by subtracting the sum of the mass of water vapor condensed during refrigeration and the defrost period from the total mass of collected condensate during total refrigeration run time (Equation 7).

EQUATION 7. MASS OF MELTED FROST DURING DEFROST PERIOD

defcondrefrigcondcondtotalfrost mmmm

where,

frostm

= Mass of melted frost during defrost, (lb/hr)

condtotalm

= Total mass of condensate collected at the end of 24-hour test, (lb/hr)

refrigcondm

= Mass of water condensed from air during refrigeration period, (lb/hr)

SENSIBLE AND LATENT LOADS After the mass flow rate of air was determined, the sensible load was calculated using Equation 8.

EQUATION 8. SENSIBLE LOAD OF REFRIGERATION

outairinairpairrefsensible TTCmQ air

where,

refsensibleQ

= Sensible load of refrigeration, (Btu/hr)

airpC = Specific heat of air, (Btu/lb-oF)

inairT = Temperature of entering air at the evaporator coil, (oF)

outairT = Temperature of existing air at the evaporator coil, (oF)



The latent load, on the other hand, was obtained by subtracting the sensible load from the total refrigeration load (Equation 9).

EQUATION 9. LATENT LOAD OF REFRIGERATION

refsensiblerefcasereflatent QQQ

where,

reflatentQ

= Latent load of refrigeration, (Btu/hr)

COOLING LOAD BASED ON ONE RUNNING CYCLE Based on ASHRAE Standard 72-05, the cooling load of the display case must be determined from one run cycle of data within the test. A running cycle refers to the refrigeration period between two defrost periods. This calculation is primarily based on refrigerant properties during the last three quarters of the running cycle. Equation 10 was used to calculate the cooling load during the last three quarters of the running cycle.

EQUATION 10. COOLING LOAD DURING THE LAST THREE-QUARTERS OF THE REFRIGERATION RUN CYCLE

cyclerunning

cyclerunningliqvapcyclerunning

tmhhQ

where,

cyclerunningQ

= Average cooling load for the running cycle, (Btu/hr)

vaph = Enthalpy of leaving refrigerant vapor during the last ¾ of the running cycle, (Btu/lb)

liqh = Enthalpy of entering liquid refrigerant during the entire running cycle, (Btu/lb)

cyclerunningm = Total refrigerant mass flow for the running cycle, (lb)

cyclerunningt = Refrigeration time period for the running cycle, (hrs)

The reduction factor is the ratio of refrigeration time period for the running cycle to overall time for one running cycle plus one defrost period (Equation 11). Multiplying the resulting reduction factor by the average cooling load for the running cycle is a reduced average cooling load for the overall time period (Equation 12).

EQUATION 11. REDUCTION FACTOR FOR REFRIGERATION RUN CYCLE

cycleoverall

cyclerunning

ttRF

where,

RF = Reduction factor, (unit-less) cycleoverallt = Overall time for one running cycle plus one defrost period, (hrs)



EQUATION 12. COOLING LOAD FOR ONE REFRIGERATION RUN CYCLE

RFQQ cyclerunningcycleoverall

where,

overall cycleQ

= Reduced average cooling load for the overall time period, (Btu/hr)

EVAPORATOR COIL CHARACTERISTIC PERFORMANCE One indication of coil performance is the temperature differential across the evaporator coil. The temperature differential across the evaporator coil was determined based on measured air temperatures at the inlet and outlet of the evaporator coil, Equation 13.

EQUATION 13. TEMPERATURE DIFFERENTIAL (T) ACROSS THE EVAPORATOR COIL

outairinairevap T TΔT

where,

evapΔT = Temperature differential across the evaporator coil, (oF)

Another indication of coil performance is the evaporator temperature difference (TD). It is defined as the difference in temperature between the temperature of the air leaving the evaporator and the saturation temperature of the refrigerant corresponding to the pressure at the evaporator coil outlet (Equation 14).

EQUATION 14. TEMPERATURE DIFFERENCE (TD) ACROSS THE EVAPORATOR COIL

SET outairevap TTD

where,

evapTD = Temperature difference across the evaporator coil, (oF)

SET = Saturated evaporator temperature based on evaporator coil outlet pressure, (oF)

The evaporator coil superheat, which was one of the system parameters, was determined as well. This parameter was obtained based on vapor refrigerant temperature at the outlet of the evaporator coil and the saturation temperature of the refrigerant corresponding to the pressure at the outlet of the evaporator coil (Equation 15).

EQUATION 15. EVAPORATOR COIL SUPERHEAT

SET TSH vapevap

where,

evapSH = Evaporator coil superheat, (oF)

Tvap = Vapor refrigerant temperature at the outlet of the evaporator coil, (oF)



Another important indication of coil performance is the ability of the coil to remove moisture from the air. This value is determined by multiplying the mass flow rate of air through the coil by the difference between the air’s absolute humidity at the coil inlet and outlet (Equation 16).

EQUATION 16. EVAPORATOR COIL MOISTURE REMOVAL RATE

kωωmMRR outairinairair

where,

MRR = Moisture removal rate of the evaporator, (lb/hr)


The evaporator heat exchange effectiveness is dependent on its log-mean temperature difference (LMTD) and its effective overall heat transfer coefficient, UA. The LMTD is determined using the refrigerant and air temperatures at the inlet and outlet of the evaporator coil according to Equation 17.

EQUATION 17. EVAPORATOR COIL LOG-MEAN TEMPERATURE DIFFERENCE (LMTD)

SETTSETTln

TTLMTD

out‐air

in‐air

out‐airin‐air

where,

LMTD = Evaporator coil log-mean temperature difference, oF

After the evaporator coil LMTD was determined, the effective overall heat transfer coefficient, UA, of the coil can be determined by the ratio of total refrigeration load to the coil LMTD (Equation 18). The UA of the evaporator coil is a function of coil material and its effective surface area.

EQUATION 18. EVAPORATOR COIL EFFECTIVE OVERALL HEAT TRANSFER COEFFICIENT (UA)

LMTDQUA refcase

where,

UA = Effective overall heat transfer coefficient of the coil, (Btu/hr-oF)

TOTAL SYSTEM POWER AND ENERGY Total system power and energy use for the tests excluded condenser power. The total system power of the fixture was obtained using Equation 19. The power usage associated with the evaporator and auxiliary or ambient fan motors, lighting system, and compressor was read directly from the data acquisition system.

EQUATION 19. TOTAL REFRIGERATION POWER USAGE, EXCLUDING CONDENSER

CompCaseLightsansSecondaryFEvapFansTotal kWkWkWkWkW +++=

where,



TotalkW = Power usage by the refrigeration system, excluding condenser, (kW)

EvapFanskW = Power usage by the evaporator fan motors, (kW)

ansSecondaryFkW = Power usage by the secondary fan motors, if applicable, (kW)

CaseLightskW = Power usage by the light fixtures in the case, (kW)

CompkW = Power usage by the compressor, (kW)

The energy consumption of the lights, evaporator fan motors, secondary or auxiliary fan motors, and the compressor is defined as the product of supplied power and total hours of power usage. Lights and evaporator fan motors stayed on continuously; hence, their total hours of power usage was equal to the total test hours (Equation 20 and Equation 22). Similarly, for the display case that was equipped with a secondary fan system, the fans operated continuously and their total hours of power usage was equal to the total test hours (Equation 21). The compressor run time, however, was a function of frequency and duration of defrost periods. The energy consumed by the compressor was determined using Equation 23.

EQUATION 20. ENERGY USAGE BY THE EVAPORATOR FAN MOTORS

EvapFansEvapFansEvapFans tkWkWh ×=

where,

EvapFanskWh = Energy consumed by the evaporator fan motors, (kWh)

EvapFanst = Total time of power usage by the evaporator fan motors, (hours)

EQUATION 21. ENERGY USAGE BY THE SECONDARY FAN MOTORS

ansSecondaryFansSecondaryFansSecondaryF tkWkWh ×=

where,

EvapFanskWh = Energy consumed by the secondary fan motors, (kWh)

EvapFanst = Total time of power usage by the secondary fan motors, (hours)

EQUATION 22. ENERGY USAGE BY THE LIGHT FIXTURES IN THE DISPLAY CASE

CaseLightsCaseLightsCaseLights tkWkWh ×=

where,

CaseLightskWh = Energy consumed by the light fixtures in the case, (kWh)

CaseLightst = Total time of power usage by the light fixtures in the case, (hours)



EQUATION 23. ENERGY USAGE BY THE COMPRESSOR

CompCompComp tkWkWh ×=

where,

CompkWh = Energy consumed by the compressor, (kWh)

Compt = Total time of power usage by the compressor, (hours)

After energy consumed by each individual component was determined, the total energy consumption for the display case was obtained using Equation 24.

EQUATION 24. TOTAL REFRIGERATION ENERGY USAGE, EXCLUDING CONDENSER

CompCaseLightsansSecondaryFEvapFansTotal kWhkWhkWhkWhkWh +++=

where,

TotalkWh = Energy usage by the refrigeration system, excluding condenser, (kWh)

DISPLAY CASE HEAT TRANSFER ANALYSIS The heat transfer within a display case involves interactions between the product and the internal environment of the case, as well as incoming heat from the surroundings into the case. The constituents of incoming heat from the surrounding environment include transmission (or conduction), infiltration and radiation. The heat from the internal sources include case lighting and evaporator fan motor(s).

Conduction and radiation loads depend on the temperatures within the case and that of ambient air. Open display cases rely on the effectiveness of their air curtains to prevent the penetration of warm and moist ambient air into the cold environment inside the case. The air curtain plays a significant role in the thermal interaction of a vertical display case and surrounding ambient air. The following sections provide a detailed discussion of the display case cooling load components, as well as methodologies employed in this project to quantify them.

TRANSMISSION (OR CONDUCTION) LOAD The transmission load refers to the conduction of heat through the display case shell. The temperature difference between the air in the room and the inside surfaces of the case is the driving force for this transfer of heat. The first task in determining the transmission load was to determine the overall coefficient of heat transfer of the case walls. This involves determining all outside and inside air film convective coefficients, thermal conductivity of the outer and inner walls of the case, and thermal conductivity of the insulation between the inner and outer walls. A simplified schematic of the display case wall assembly layers is shown in Figure 10. Equation 25 describes the approach used to determine the overall coefficient of heat transfer for the display case.



Inner Shell of Case

Insulation Between the Inner and OuterShell of Case

Outer Shell of Case

FIGURE 10. SCHEMATICS OF INNER AND OUTER SHELL OF THE CASE AND INSULATION BETWEEN THEM

EQUATION 25. OVERALL HEAT TRANSFER COEFFICIENT FOR THE DISPLAY CASE WALLS

o3

3

2

2

1

1

i h1

kL

kL

kL

h1

1U

where,

U = Overall coefficient of heat transfer for the case walls, (Btu/hr-ft2-F)

ih = Convective coefficient for inside case air film against case inner wall, (Btu/hr-ft2-F)

1L = Thickness of outer shell of the case, (in) 1k = Thermal conductivity of outer shell of case, (Btu-in/hr-ft2-F)

2L = Thickness of insulation within the case walls, (in) 2k = Thermal conductivity of insulation within the case walls,

(Btu-in/hr-ft2-F)

3L = Thickness of inner shell of the case, (in)

3k = Thermal conductivity of inner shell of case, (Btu-in/hr-ft2-F)

oh = Convective coefficient for outside/room air film against case outer shell, (Btu/hr-ft2-F)

After the overall coefficient of heat transfer was determined, the transmission load was determined using Equation 26. The inside temperature of various surfaces inside the case was assumed to be in equilibrium with the air temperature inside the case.

EQUATION 26. TRANSMISSION OR CONDUCTION LOAD OF THE DISPLAY CASE

)T(TAUQ caseroomcond

where,

condQ

= Transmission, or conduction, load of the case, (Btu/hr)



A = Total surface area of case walls that are conducting heat, (ft2) roomT = Dry bulb temperature of the air in the room, (F)

caseT = Dry bulb temperature of the air inside the case, (F)

RADIATION LOAD The temperature of walls inside the controlled environment room was assumed to be equal to the temperature of air inside the room. Similar to the conduction analysis, the inside temperature of various surfaces inside the case was assumed to be equal to the air temperature inside the case. This assumption was later verified and accepted after the subject temperatures were measured individually and were found to be equal to the air temperature adjacent to them. The case load, due to radiation heat transfer, was determined by simply modeling the system as two gray surfaces, one surface representing the total surface area of the room (walls, floor, ceiling), and the other being an imaginary plane covering the opening of the display case. All of the radiation leaving the room surfaces will arrive at the imaginary plane. The imaginary plane at the case opening will, in turn, exchange all of its radiation with the interior surfaces of the display case. A series of calculations were performed to develop the effective view factor between the room and inside of the case using Kirchoff’s Law and the reciprocity relation.

Figure 11 shows a simplified plan view of the controlled environment room and the surfaces exchanging heat through radiation with the display case. The surfaces inside of the display case (back, top, bottom, and sides) were all designated as surface 1, the room surfaces were designated as surface 2, and the imaginary plane covering the case opening was designated as surface 3. From the reciprocity relation, A1F1-3 = A3F3-1. In this case, F3-1 is 1, and F1-3 = F1-2, therefore, F1-2 = A3/A1. After this view factor was determined, Equation 27 was used to calculate the radiation load of the cases.



Surface 1(Case Interior)

T11A1

Surface 3(Imag. Plane)

T33A3

Surface 2(Room)

T22A2

FIGURE 11. SURFACES PARTICIPATING IN DISPLAY CASE RADIATION HEAT TRANSFER

EQUATION 27. RADIATION LOAD OF THE DISPLAY CASE

cc

c

cwwww

w

4c

4wrad

Aεε1

FA1

Aεε1

TTσQ

where,

radQ

= Radiation heat transfer between room walls and display case, (Btu/hr)

= Stefan-Boltzmann Constant, (0.1714 * 10-8 Btu/hr-ft2-R4)

wT = Surface temperature of the room walls, (R)

cT = Surface temperature of the display case inner walls, (R)

w = Emissivity of the room walls

wA = Total area of room surfaces, (ft2)

cwF = View factor from case to surfaces of the room c = Emissivity of the inside walls of the case

cA = Total area of the inside walls of the case, (ft2)

INTERNAL LOAD The internal load of the display case refers to the heat introduced and dissipated by its internal components. The internal load for the display cases under consideration includes the heat introduced by the case lighting and by the evaporator fan motors. The fan motors, lamps, and ballasts are located inside the thermodynamic boundary of each case. Hence, their total heat dissipation was considered part of the case load.



For the display case with secondary fan assembly, the heat dissipated from the fans was not considered part of the case load since they were located outside the thermodynamic boundary of the case. The power consumed by these devices was recorded directly by the data logger, which was then converted to a cooling load according to Equation 28 and Equation 29.

EQUATION 28. DISPLAY CASE LOAD DUE TO EVAPORATOR FAN MOTORS

KkWQ EvapFansEvapFans

where,

EvapFansQ

= Case load due to fan motors, (Btu/hr)

EvapFanskW = Power consumed by the evaporator fan motors, (kW)

K = Conversion factor, (3,413 Btu/hr/kW)

EQUATION 29. DISPLAY CASE LOAD DUE TO LIGHTING

KkWQ CaseLightsCaseLights

where,

CaseLightsQ

= Case load due to lighting, (Btu/hr)

CaseLightskW = Power consumed by the light fixtures in the case, (kW)

K = Conversion factor, (3,413 Btu/hr/kW)

INFILTRATION LOAD The infiltration load of the display case refers to the entrainment of warm and moist air from the room, across the case air curtain, into the refrigerated space. The infiltration load has two components—sensible and latent. The sensible portion refers to the temperature-driven heat penetrating into the display case, whereas t

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