hussmann display case

Transcription

Design & Engineering Services
Performance Comparison of Three High
Efficiency Medium-Temperature Display Cases
ET 06-07
Prepared by:
Customer Service Business Unit
Southern California Edison
June 22, 2009
Performance Comparison of Three High Efficiency Medium-Temperature Display Cases
ET 06.07
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 Rafik.Sarhadian@sce.com.
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.
March 2009
ET 06.07
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
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
ET 06.07
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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
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
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
Figure 34. Two-minute Profile of Product Temperature at Six
Different Locations for Bottom Shelf Over 24 Hours –
Different Locations for Second Shelf Over 24 Hours –
Different Locations for Third Shelf Over 24 Hours – Hill
Different Locations for Fourth Shelf Over 24 Hours – Hill
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Different Locations for Top Shelf Over 24 Hours – Hill
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
Dry Bulb and Relative Humidity Over 24 Hours –
Hussmann Display Case ............................................. 49
Over 24 Hours – Hussmann Display Case ...................... 49
Temperatures Over 24 Hours – Hussmann Display case .. 50
Hours – Hussmann Display Case .................................. 50
Relative Humidity Over 24 Hours – Hussmann Display
Case ........................................................................ 51
the Display Case Over 24 Hours – Hussmann Display
Case ........................................................................ 52
Figure 48. Cooling Load by Component Over 24 Hours – Hussmann
Display Case ............................................................. 52
Over 24 hours – Hussmann Display Case ...................... 53
24 Hours and ¾ of Running Cycle – Hussmann Display
Case ........................................................................ 53
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
Difference (TD) Over 24 Hours – Hussmann Display
Case ........................................................................ 55
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 –
Different Locations for Third Shelf Over 24 Hours –
Different Locations for Top Shelf Over 24 Hours –
Hours – Hussmann Display Case .................................. 59
Over 24 Hours – Hussmann Display Case ...................... 59
Dry Bulb and Relative Humidity Over 24 Hours – Tyler
Display Case ............................................................. 61
Over 24 Hours – Tyler Display Case ............................. 61
Temperatures Over 24 Hours – Tyler Display case .......... 62
Figure 66. Individual and Average Discharge Air Temperature Over
24 Hours – Tyler Display case ..................................... 62
Hours – Tyler Display Case ......................................... 63
Tyler Display Case ..................................................... 63
Relative Humidity Over 24 Hours – Tyler Display Case .... 64
the Display Case Over 24 Hours – Tyler Display Case ..... 64
Figure 71. Cooling Load by Component Over 24 Hours – Tyler
Display Case ............................................................. 65
Over 24 hours – Tyler Display Case.............................. 65
24 Hours and ¾ of Running Cycle – Tyler Display Case ... 66
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
Difference (TD) Over 24 Hours – Tyler Display Case ....... 67
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 –
Different Locations for Third Shelf Over 24 Hours –
Different Locations for Fourth Shelf Over 24 Hours –
Different Locations for Top Shelf Over 24 Hours – Tyler
Display Case ............................................................. 71
Hours – Tyler Display Case ......................................... 72
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
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 –
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
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 –
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.
Published Data – Hussmann Display Case ..................... 60
Table 6.
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
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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
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June 2009
ET 06.07
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)
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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 energyintensive 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 multideck [Refs. 1, 2].
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June 2009
Medium -Tem p
Island Cases
Medium -Tem p
11%
Single-Deck
Open Cases
3%
Medium -Tem p
Multi-Deck
Open Vertical
50%
FIGURE 1.
Medium -Tem p
Service Cases
4%
ET 06.07
Low Tem p
Reach-ins
33%
Medium -Tem p
Reach-ins
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.
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TABLE 1.
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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 recirculated 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.
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June 2009
FIGURE 2.
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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].
Page 6
June 2009
Evaporator
Fans
3%
Case
Lighting
6%
Conduction
3%
ET 06.07
Radiation
8%
Infiltration
81%
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.
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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.
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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 AirConditioning Engineers (ASHRAE) Standard 72-05, which requires the lighting intensity not
be less than 75 foot-candles at this location.
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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
Open 4-Deck
Front Loading
Hussmann
M5X-GEP
Meat and Deli
(MT)
30.0
1,370
8
Open 5-Deck
Front Loading
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 footcandles. 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.
Page 10
June 2009
FIGURE 5.
ET 06.07
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.
product simulator FIGURE 6.
dummy products 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.
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MAIN CHANNELS DESCRIPTION
FIGURE 7.
E
FRONT PRODUCT TEMPERATURES
F
BACK PRODUCT TEMPERATURES
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  0.01C

Analog inputs from pressure transducers
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June 2009

Dew point sensors

Flow meters

CT-transducers
ET 06.07
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 10second 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 PC5062BX680
+ 0.5% F.S. (0.04 kW)
Power
Ohio Semitronics Model P-143B
+ 1.0% F.S. (0.08 kW)
Pressure
Setra Transducers Model C207100 & 500 PSIG Pressure
Ranges
+ 0.13%
Pressure
Danfoss Transducers Model
AKS32
0-500 PSIG
+ 0.2% F.S.
Temperature (RTD)
Hy-Cal Engineering Model RTS37-A-100
+ 0.01oC
Temperature (TC)
Kaye Instruments T/W 50
through 80; Melt # 8032
+ 0.1oC
Scale
HP-30K
+ 0.1 gram
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June 2009
ET 06.07
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)
FIGURE 8.
i
LEFT
j
MIDDLE
k
RIGHT
ik
BETWEEN RIGHT AND MIDDLE
jk
BETWEEN LEFT AND MIDDLE
ijk
COMBINATION OF RIGHT, MIDDLE, AND LEFT
LOCATION OF SENSORS FOR OPEN VERTICAL MULTI-DECK DISPLAY CASES
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June 2009
ET 06.07
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 10second 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 10second 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
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June 2009
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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.
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June 2009
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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)

Q caseref  m ref  RE  k
where,

Q caseref

= Total refrigeration load of the case, sensible and latent, (Btu/hr)
m ref
= 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)

Q caseref
Q caseref (tons ) 
12,000

where,

Q caseref (tons )
= 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
Page 17
June 2009
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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
cfmcase   Aback  panel  ADAG   DAVavg
where,
cfmcase
= Volumetric flow rate of air into the display case, (ft3/min)
Aback panel
= Total area of openings in the back panel, (ft2)
ADAG
= Total free area available through discharge air grille, (ft2)
DAVavg
= 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

m air  cfmcase  airin  k
where,

m air
= Mass flow rate of air, (lb/hr)
airin
= Density of air at the inlet of the evaporator coil, (lb/ft3)
k
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.

m conddef
MASS OF CONDENSATE COLLECTED FROM AIR DURING DEFROST PERIOD
 airin- airout   m air  tdefrost 



trefrig
where,
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June 2009

ET 06.07
m conddef
= Mass of water vapor condensed from air during defrost period, (lb/hr)
airin
= Absolute humidity of air at the evaporator inlet, (lbw/lba)
airout
= Absolute humidity of air at the evaporator outlet, (lbw/lba)
tdefrost
= Defrost period, (hours)
trefrig
= 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




m frost  m totalcond   m condrefrig  m conddef 


where,

m frost

m totalcond
= Mass of melted frost during defrost, (lb/hr)
= Total mass of condensate collected at the end of 24-hour test, (lb/hr)

m cond refrig = 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
Q sensibleref  m air  Cpair  Tairin  Tairout 


where,

Q sensibleref = Sensible load of refrigeration, (Btu/hr)
Cpair
= Specific heat of air, (Btu/lb-oF)
Tairin
= Temperature of entering air at the evaporator coil, (oF)
Tairout
= Temperature of existing air at the evaporator coil, (oF)
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June 2009
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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


Q latentref  Q caseref  Q sensibleref
where,

Q latentref
= 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
Q runningcycle 
hvap  hliq   mrunningcycle
trunningcycle
where,

Q runningcycle = Average cooling load for the running cycle, (Btu/hr)
hvap
= Enthalpy of leaving refrigerant vapor during the last ¾ of the running
cycle, (Btu/lb)
hliq
= Enthalpy of entering liquid refrigerant during the entire running cycle,
(Btu/lb)
mrunningcycle = Total refrigerant mass flow for the running cycle, (lb)
trunningcycle = 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.
RF 
REDUCTION FACTOR FOR REFRIGERATION RUN CYCLE
trunningcycle
toverallcycle
where,
RF
= Reduction factor, (unit-less)
toverallcycle
= Overall time for one running cycle plus one defrost period, (hrs)
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June 2009
EQUATION 12.

ET 06.07
COOLING LOAD FOR ONE REFRIGERATION RUN CYCLE

Q overallcycle  Q runningcycle  RF
where,

Q overallcycle = 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
ΔTevap  Tairin  Tairout
where,
ΔTevap
= 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
TDevap  Tairout  SET
where,
TDevap
= 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
SHevap  Tvap  SET
where,
SHevap
= Evaporator coil superheat, (oF)
Tvap = Vapor refrigerant temperature at the outlet of the evaporator coil, (oF)
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June 2009
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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
MRR  m air  ωairin  ωairout   k

where,
MRR
= Moisture removal rate of the evaporator, (lb/hr)
k
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.
LMTD 
EVAPORATOR COIL LOG-MEAN TEMPERATURE DIFFERENCE (LMTD)
Tair‐in  Tair‐out 
  Tair‐in  SET 
ln Tair‐out  SET 

 
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)

UA 
Q caseref
LMTD
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
kWTotal = kWEvapFans + kWSecondaryFans + kWCaseLights + kWComp
where,
Page 22
June 2009
ET 06.07
kWTotal
= Power usage by the refrigeration system, excluding condenser, (kW)
kWEvapFans
= Power usage by the evaporator fan motors, (kW)
kWSecondaryFans
= Power usage by the secondary fan motors, if applicable, (kW)
kWCaseLights
= Power usage by the light fixtures in the case, (kW)
kWComp
= 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
kWhEvapFans = kWEvapFans × tEvapFans
where,
kWhEvapFans = Energy consumed by the evaporator fan motors, (kWh)
tEvapFans
EQUATION 21.
= Total time of power usage by the evaporator fan motors, (hours)
ENERGY USAGE BY THE SECONDARY FAN MOTORS
kWhSecondaryFans = kWSecondaryFans × tSecondaryFans
where,
kWhEvapFans = Energy consumed by the secondary fan motors, (kWh)
tEvapFans
EQUATION 22.
= Total time of power usage by the secondary fan motors, (hours)
ENERGY USAGE BY THE LIGHT FIXTURES IN THE DISPLAY CASE
kWhCaseLights = kWCaseLights × tCaseLights
where,
kWhCaseLights = Energy consumed by the light fixtures in the case, (kWh)
tCaseLights
= Total time of power usage by the light fixtures in the case, (hours)
Page 23
June 2009
EQUATION 23.
ET 06.07
ENERGY USAGE BY THE COMPRESSOR
kWhComp = kWComp × tComp
where,
kWhComp
= Energy consumed by the compressor, (kWh)
tComp
= 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
kWhTotal = kWhEvapFans + kWhSecondaryFans + kWhCaseLights + kWhComp
where,
kWhTotal
= 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.
Page 24
June 2009
ET 06.07
Inner Shell of Case
Insulation Between the Inner and Outer
Shell of Case
Outer Shell of Case
FIGURE 10. SCHEMATICS OF INNER AND OUTER SHELL OF THE CASE AND INSULATION BETWEEN THEM
EQUATION 25.
U
OVERALL HEAT TRANSFER COEFFICIENT FOR THE DISPLAY CASE WALLS
1
 1   L1   L2   L3   1 
 hi    k1    k2    k3    ho 
         
where,
U
= Overall coefficient of heat transfer for the case walls, (Btu/hr-ft2-F)
hi
= Convective coefficient for inside case air film against case inner wall,
(Btu/hr-ft2-F)
L1
= Thickness of outer shell of the case, (in)
k1
= Thermal conductivity of outer shell of case, (Btu-in/hr-ft2-F)
L2
= Thickness of insulation within the case walls, (in)
k2
= Thermal conductivity of insulation within the case walls,
(Btu-in/hr-ft2-F)
L3
= Thickness of inner shell of the case, (in)
k3
= Thermal conductivity of inner shell of case, (Btu-in/hr-ft2-F)
ho
= 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

Q cond  U  A  (Troom  Tcase)
where,

Q cond
= Transmission, or conduction, load of the case, (Btu/hr)
Page 25
June 2009
A
= Total surface area of case walls that are conducting heat, (ft2)
Troom
= Dry bulb temperature of the air in the room, (F)
Tcase
= Dry bulb temperature of the air inside the case, (F)
ET 06.07
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.
Page 26
June 2009
ET 06.07
Surface 3
(Imag. Plane)
T3
3
A3
Surface 2
(Room)
T2
2
A2
Surface 1
(Case Interior)
T1
1
A1
FIGURE 11. SURFACES PARTICIPATING IN DISPLAY CASE RADIATION HEAT TRANSFER
EQUATION 27.
RADIATION LOAD OF THE DISPLAY CASE


σ  Tw  Tc
Q rad 
 1  εw  
1
  1  εc  
 εw  Aw    Aw  Fcw    εc  Ac 
 
 



4
4
where,

Q rad
= Radiation heat transfer between room walls and display case, (Btu/hr)

= Stefan-Boltzmann Constant, (0.1714 * 10-8 Btu/hr-ft2-R4)
Tw
= Surface temperature of the room walls, (R)
Tc
= Surface temperature of the display case inner walls, (R)
w
= Emissivity of the room walls
Aw
= Total area of room surfaces, (ft2)
Fcw
= View factor from case to surfaces of the room
c
= Emissivity of the inside walls of the case
Ac
= 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.
Page 27
June 2009
ET 06.07
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

Q EvapFans  kWEvapFans  K
where,

Q EvapFans
= Case load due to fan motors, (Btu/hr)
kWEvapFans = Power consumed by the evaporator fan motors, (kW)
K
EQUATION 29.
= Conversion factor, (3,413 Btu/hr/kW)
DISPLAY CASE LOAD DUE TO LIGHTING

Q CaseLights  kWCaseLights  K
where,

Q CaseLights
= Case load due to lighting, (Btu/hr)
kWCaseLights = 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 the latent
portion refers to the heat content of moisture within the infiltrating air. As air passes
through the evaporator, it loses its sensible heat and dehumidifies.
A reverse calculation approach was used to determine the infiltration load of the
display cases. After obtaining the total case load along with all other cooling
components, Equation 30 was used to obtain the total infiltration load.
EQUATION 30.
INFILTRATION LOAD OF THE DISPLAY CASE







Q inf  Q caseref   Q EvapFans  Q CaseLights  Q cond  Qrad 


where,

Q inf

Q caseref
= Total load added to the case due to infiltration of room air, (Btu/hr)
= Total refrigeration load of the case determined by refrigerant properties,
(Btu/hr)
Page 28
June 2009
ET 06.07
This approach relies on a mass energy balance solution, which cannot be directly
influenced by the airflow characteristics of the air curtain. Consequently, the effects
of discharge air velocity, discharge grille dimensions, and other geometry related
characteristics of each case did not play a direct role in determining the mass of
warm and moist air entrained into the cases. The flow rate of air into the display
case was determined using Equation 31.
EQUATION 31.
cfminf 
VOLUMETRIC FLOW RATE OF INFILTRATED AIR FROM ROOM INTO THE DISPLAY CASE
mtotal‐cond  mcond‐def 
ωair‐room  ωair‐case   ρair‐room  trefrig 
where,
cfminf
= Amount of infiltrated air from the room into the display case, (ft3/min)
mtotalcond
= Total mass of condensate collected over 24-hour test, (lb)
mconddef
= Total mass of water vapor condensed from air during defrost periods,
(lb)
ωair‐room
= Absolute humidity of air in the room, (lbw/lba)
ωair‐case
= Absolute humidity of air in the case, (lbw/lba)
ρair‐room
= Density of air in the room, (lb/ft3)
trefrig
= Refrigeration period, (minutes)
Additionally, the sensible and latent load components of the total infiltration load
were obtained. The sensible portion of the infiltration load was determined using
Equation 32.
EQUATION 32.
SENSIBLE PORTION OF THE INFILTRATION LOAD OF THE DISPLAY CASE
Q sensible‐inf  cmfinf  ρair  Cpair  Troom  Tcase   k

where,

Q sensible‐inf
= Sensible part of the infiltration load, (Btu/hr)
ρair
= Density of air, lb/ft3
Cpair
= Specific heat of air, (Btu/lb)
k
The latent portion of the infiltration, however, was obtained by subtracting the
sensible load from the total infiltration load as shown in Equation 33.
EQUATION 33.

LATENT PORTION OF THE INFILTRATION LOAD OF THE DISPLAY CASE


Q latentinf  Q inf  Q sensibleinf
where,

Q latent‐inf
= Latent part of the infiltration load, (Btu/hr)
Page 29
June 2009
ET 06.07
DESCRIPTION OF DISPLAY CASES
This section describes all three tested new-generation high efficiency MT OVRDCs. All three
acquired cases were standard high efficiency models without any extra options or features.
The display case manufacturers and their corresponding model numbers are:
1. Hill Phoenix – O5DM
2. Hussmann – M5X-GEP
3. Tyler – N6DHPACLA
HILL PHOENIX DISPLAY CASE – O5DM
The following are the specifications for the Hill Phoenix eight-foot five-deck display
case tested in this project. Figure 12 and Figure 13 depict the photograph and the
schematic diagram of the case with all the important dimensions shown. As a
standard feature and integral part of the case, the O5DM deli case was manufactured
with two-and-a-half inch extended front or front sill height without nose light.
Evaporator:
Evaporator fan motor:
Evaporator fan blade:
Air curtain:
Honeycomb:
Number of Shelves:
Expansion valve:
Defrost type:
Defrost frequency:
Defrost length:
Defrost termination temp:
Refrigerated volume:
One coil per case
8.66” deep x 7.5” tall x 129” wide
6 circuits
8 tubes per circuit, smooth copper tube,
0.016” tube wall thickness
0.375” tube nominal outside diameter
Corrugated fins, 0.0075” fin thickness, 4 fins/inch
Three high-efficiency fans (ECM), 9 watt
8” diameter, 5 blades, 37° pitch
Single band
4” wide, 1” deep, 1/8” holes
Five
Sporlan ESX electronic expansion valve
Off-cycle
Four times per day
42 minutes (fail-safe)
47oF
92.11 ft3
Refrigeration Data
Refrigerant:
Discharge air:
Discharge air velocity:
Return air:
Evaporator:
Conventional capacity:
Superheat set point:
R-404A
30oF
270 fpm
44oF
22oF
1,570 BTUH @ 22oF
6-8oF
Electrical Data
Fans:
Lighting:
120 volts, 0.70 amps (high-efficiency fans, ECM)
120 volts, 0.47 amps per light row (two 4-foot
T8s with electronic ballast per light row)
Page 30
June 2009
ET 06.07
FIGURE 12. PHOTOGRAPH OF HILL PHOENIX’S 8-FOOT, 5-DECK DISPLAY CASE
FIGURE 13. SCHEMATIC OF THE 8-FOOT, 5-DECK DISPLAY HILL PHOENIX CASE (COURTESY OF HILL PHOENIX)
Page 31
June 2009
ET 06.07
HUSSMANN DISPLAY CASE – M5X-GEP
The following are the specifications for the Hussmann’s eight-foot four-deck display
case tested in this project. Figure 14 and Figure 15 depict the photograph and the
standard feature, the M5X-GEP deli/meat case was equipped with a seven-and-a-half
inch glass-front extension.
Evaporator:
Evaporator fan motor:
Evaporator fan blade:
Air curtain:
Honeycomb:
Number of Shelves:
Expansion valve:
Defrost type:
Defrost frequency:
Defrost length:
Defrost termination temp:
60 pass, 42.5” tubes
4.800” x 21.150”
Flat fins, 4.5 fins/inch
39.333” finned length
Two coils per case
2 circuits
Tube on tube
Four high efficiency fans (ECM: 72 watts)
10” diameter, 30° pitch
Single band
5” wide, 1” to 2” tapered thick, 0.157” dia. cell
Four
TXV-R404A Hussmann TD1 SWT
Off-cycle
Four times per day
48oF
83.92 ft3
Refrigeration Data
Refrigerant:
Discharge air:
Evaporator:
R-404A
30oF
26oF
1,380 BTUH @ 26oF, Lit
Electrical Data
Fans:
Lighting:
F32 T8s with electronic ballast per light row)
Page 32
June 2009
ET 06.07
FIGURE 14. PHOTOGRAPH OF HUSSMANN’S 8-FOOT, 4-DECK DISPLAY CASE
FIGURE 15. SCHEMATIC OF THE 8-FOOT, 4-DECK HUSSMANN DISPLAY CASE (COURTESY OF HUSSMANN)
Page 33
June 2009
ET 06.07
TYLER DISPLAY CASE – N6DHPACLA
The following are the specifications for the Tyler eight-foot five-deck display case
tested in this project. Figure 16 and Figure 17 depict the photograph and the
standard feature, this display case was not equipped with any front extension.
Evaporator:
Evaporator (primary) fan:
Secondary Air Curtain fan:
Air curtain:
Honeycomb:
Number of Shelves:
Defrost type:
Defrost frequency:
Defrost length:
One coil per case
9.5” deep x 6.25” tall x 80” wide
4 circuits
8 tubes per circuit, smooth copper tube
0.016” tube wall thickness
0.375” tube nominal outside diameter
Corrugated fins, 0.0095” fin thickness, 6 fins/inch
Two high-efficiency fans (ECM: 34 watts)
Two high-efficiency fans (ECM: 22 watts)
Dual band
9” wide, 1” deep, 1/8” holes
Five
Off-cycle
Six times per day
93.00 ft3 (estimated)
Refrigeration Data
Refrigerant:
Discharge air:
Discharge air velocity:
Evaporator:
R-404A
34.5oF
110 fpm (primary/evaporator fans)
28oF
1,059 Btu/h @ 28oF, Lit
Electrical Data
Fans:
Lighting:
T8s with electronic ballast per light row)
Page 34
June 2009
ET 06.07
FIGURE 16. PHOTOGRAPH OF TYLER’S 8-FOOT, 5-DECK DISPLAY CASE
FIGURE 17. SCHEMATIC OF THE 8-FOOT, 5-DECK TYLER DISPLAY CASE (COURTESY OF TYLER)
Page 35
June 2009
ET 06.07
RESULTS
The thermal analysis for all three new-generation high efficiency MT OVRDCs was
performed. The analysis was conducted in accordance with manufacturers’ specified DAT. In
addition, all tests were performed under American Society of Heating, Refrigerating and Airconditioning Engineers (ASHRAE) Standard 72-05. Subsequent sections provide thermal test
results for all three display cases.
HILL PHOENIX DISPLAY CASE (O5DM)
80
75
70
65
60
55
50
45
40
35
30
25
20
85
65
55
45
35
25
Room RH (%)
75
15
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
17:12:39
16:10:39
15:08:39
14:06:39
13:04:39
12:02:39
9:58:39
11:00:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
5
2:44:39
Room Temp (F)
The performance of the Hill Phoenix display case was evaluated under ASHRAE 72-05
conditions, 75oF DB and 55% RH. The test ran for a period of 24 hours. Prior to
initiating the test run, however, the controlled environment room was allowed to
reach a steady-state equilibrium condition. Figure 18 illustrates the two-minute
profile of the controlled environment room DB and RH during the entire test period.
As illustrated, the indoor conditions remained fairly unchanged. The average room
DB and RH was 75oF and 55.3%, respectively, which corresponded to a wet bulb
(WB) of 64oF.
Test Period (24-hours, including defrost)
Room Temp
Room RH
FIGURE 18. TWO-MINUTE PROFILE OF THE CONTROLLED ENVIRONMENT ROOM DRY BULB AND RELATIVE HUMIDITY
OVER 24 HOURS – HILL PHOENIX DISPLAY CASE
In order to maintain a DAT of 30oF or an evaporator temperature of 22oF, as
specified by the manufacturer, the test rack controller was programmed to run at a
fixed suction pressure of 58 psig (Figure 19). The rack controller was also
programmed to run at a fixed discharge pressure of 220 psig or 95oF SCT. Figure 19
illustrates the two-minute profile of suction and discharge pressures over the entire
test period.
Page 36
June 2009
ET 06.07
230
Suction Pressure (psig)
66
64
220
62
210
60
200
58
56
190
54
180
52
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
17:12:39
16:10:39
15:08:39
14:06:39
13:04:39
12:02:39
9:58:39
11:00:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
50
3:46:39
170
2:44:39
Discharge pressure (psig)
Discharge Pressure
Suction Pressure
FIGURE 19. TWO-MINUTE PROFILE OF SUCTION AND DISCHARGE PRESSURES OVER 24 HOURS – HILL PHOENIX
DISPLAY CASE
Figure 20 illustrates the two-minute profile of the average DAT and return air
temperature. As shown, the refrigeration system maintained a relatively constant
average DAT of 30oF during the entire test period. In addition, the average return air
temperature was 44.6oF, which was in close agreement with the manufacturer’s data
of 44oF.
60
Average Discharge &
Return Air Temp (F)
55
50
45
40
35
30
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
17:12:39
16:10:39
15:08:39
14:06:39
13:04:39
12:02:39
11:00:39
9:58:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
2:44:39
25
Discharge Air Temp
Return Air Temp
FIGURE 20. TWO-MINUTE PROFILE OF AVERAGE DISCHARGE AND RETURN AIR TEMPERATURES OVER 24 HOURS –
HILL PHOENIX DISPLAY CASE
Figure 21 depicts the two-minute profile of the mass of condensate collected over
the entire test period. The condensate mass was comprised of moisture collected
from the moist air stream during refrigeration run time and melted ice (or frost)
during off-cycle defrost periods. The stepped (horizontal) profiles indicate the
moisture collected during the refrigeration run time between each of the six defrost
periods. The vertical (or sloped) profiles, on the other hand, indicate the melted ice
during each defrost period.
Page 37
June 2009
ET 06.07
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
17:12:39
16:10:39
15:08:39
14:06:39
13:04:39
12:02:39
11:00:39
9:58:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
100
90
80
70
60
50
40
30
20
10
0
2:44:39
Mass of Collected
Condensate (lbs)
FIGURE 21. TWO-MINUTE PROFILE OF COLLECTED CONDENSATE OVER 24 HOURS – HILL PHOENIX DISPLAY CASE
In addition to condensate collected during refrigeration and defrosts’ ice melting,
further moisture was detected to escape from the air during defrost. During off-cycle
defrost periods, the compressor stops running while the evaporator fan motors
continue to operate, thereby bringing relatively warm and humid air into the display
case. As a result, the room’s warm and moist air was the main factor responsible for
melting the ice on the coil. Figure 22 shows the sub-components of condensation.
Clearly, the most condensate removal took place during the ice melting stages of the
defrost period (79.18 pounds).
Mass of Collected Condensate
Over 24-hours (lbs)
120
95.46
100
79.18
80
60
40
20
7.43
8.80
Mass of water vapor
condensed from air
during defrosts
Mass of water vapor
condensated from air
during refrigeration
periods
0
Mass of melted frost
during defrosts
Total measured
condensate
Breakdown of Condensate Components
FIGURE 22. BREAKDOWN OF CONDENSATE COLLECTED OVER 24 HOURS – HILL PHOENIX DISPLAY CASE
Bringing warm and humid indoor air into the case to melt frost on the coil caused the
temperature and RH inside the fixture to increase and reach maximum levels during
defrost periods (Figure 23). After the refrigeration period was initiated, the
temperature and humidity inside the case were lowered.
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June 2009
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95
85
75
65
55
45
35
25
Inside Case RH (%)
80
75
70
65
60
55
50
45
40
35
30
25
20
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
17:12:39
16:10:39
15:08:39
14:06:39
13:04:39
12:02:39
9:58:39
11:00:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
15
2:44:39
Inside Case Temp (F)
Inside Case Temp
Inside Case RH
FIGURE 23. TWO-MINUTE PROFILE OF DISPLAY CASE TEMPERATURE AND RELATIVE HUMIDITY OVER 24 HOURS –
Figure 24 depicts the total cooling load per linear foot of the case. The highest
cooling load was observed at the end of each defrost period due to bringing relatively
warm and humid air into the case during defrosts. The lowest cooling load was
observed prior to initiating defrost. The average cooling load was 2,015 Btu/hr/ft.
2,500
Cooling Load (Btu/hr/ft)
[using refrigeration data]
2,400
2,300
2,200
2,100
2,000
1,900
1,800
1,700
1,600
1,500
1,400
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Test Period (24-hours, excluding defrost)
FIGURE 24. HOURLY PROFILE OF TOTAL COOLING LOAD PER LINEAR FOOT OF THE DISPLAY CASE OVER 24
HOURS – HILL PHOENIX DISPLAY CASE
As illustrated in Figure 25 and Figure 26, the total cooling load of the display case
consisted of the infiltration, radiation, conduction, and internal loads (lights and
evaporator fans). The largest component of the cooling load was infiltration, 13,881
Btu/hr, corresponding to 86% of the total cooling load. The smallest component was
the display case’s lighting system and evaporator fan motors (internal load), 592
Btu/hr, corresponding to about 4% of the total cooling load. The conduction
accounted for roughly 4% and radiation for about 6% of the total cooling load.
Page 39
June 2009
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16,000
13,881
Cooling Load (Btu/hr)
14,000
12,000
10,000
8,000
6,000
4,000
2,000
637
1,010
Conduction
Radiation
592
0
Infiltration
Internal (lights & evap
fans)
Cooling Load Components
FIGURE 25. COOLING LOAD BY COMPONENT OVER 24 HOURS – HILL PHOENIX DISPLAY CASE
fans)
3.7%
Conduction
4.0%
Radiation
6.3%
Infiltration
86.1%
FIGURE 26. PERCENTAGE BREAKDOWN OF THE COOLING LOAD COMPONENTS OVER 24 HOURS – HILL PHOENIX
DISPLAY CASE
Additionally, the reduced cooling load, and average cooling load over the entire test
period and during the last three-fourths (3/4) of the running cycle were determined
according to ASHRAE Standard 72-05 (Figure 27). The running cycle refers to the
refrigeration period between two defrost periods.
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June 2009
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18,000
16,119
16,016
16,000
14,173
14,000
12,000
10,000
8,000
6,000
4,000
2,000
0
Average Cooling Load (over 24hours)
Average Cooling Load (3/4 of
running cycle)
Reduced Average Cooling Load
FIGURE 27. REDUCED COOLING LOAD, AND AVERAGE COOLING LOAD OVER 24 HOURS AND ¾ OF RUNNING
CYCLE – HILL PHOENIX DISPLAY CASE
The mass flow rate of refrigerant was observed to decline during each running cycle
(Figure 28). The flow rate was highest at the end of the defrost period and lowest
prior to initiation of the defrost period. This observed profile in refrigerant mass flow
rate was attributed to the change in total cooling load of the case coupled with
maintaining a fixed suction pressure during the entire test period.
8
Refrigerant Mass
Flow Rate (lb/min)
7
6
5
4
3
y = -0.0004x + 4.6902
R2 = 0.0085
2
1
2:12:39
1:22:39
0:32:39
23:42:39
22:52:39
21:30:39
20:40:39
19:50:39
19:00:39
17:38:39
16:48:39
15:58:39
15:08:39
13:46:39
12:56:39
12:06:39
9:56:39
11:16:39
9:06:39
8:16:39
7:26:39
6:06:39
5:16:39
4:26:39
3:36:39
2:46:39
0
FIGURE 28. TWO-MINUTE PROFILE OF REFRIGERANT MASS FLOW RATE OVER 24 HOURS – HILL PHOENIX
DISPLAY CASE
Comparing 2-minute compressor power and refrigerant mass flow rate profiles
revealed a close similarity in behavior between the two parameters (Figure 29), as
expected. That is, maintaining fixed suction and discharge pressure resulted in the
compressor power being entirely dependent on variations in the refrigerant mass
flow rate.
Page 41
June 2009
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6
y = -0.0004x + 4.6902
R2 = 0.0085
5
4
4
3
2
1
0
3
2
1
2:12:39
1:22:39
0:32:39
23:42:39
22:52:39
21:30:39
20:40:39
19:50:39
19:00:39
17:38:39
16:48:39
15:58:39
15:08:39
13:46:39
12:56:39
12:06:39
9:56:39
11:16:39
0
9:06:39
8:16:39
7:26:39
6:06:39
5:16:39
4:26:39
3:36:39
y = -0.0001x + 2.2505
R2 = 0.0239
Compressor Power (kW)
8
7
6
5
2:46:39
Refrigerant Mass
Flow Rate (lb/min)
Refrigerant Mass Flow Rate
Linear (Refrigerant Mass Flow Rate)
Compressor Power
Linear (Compressor Power)
FIGURE 29. TWO-MINUTE PROFILE OF COMPRESSOR POWER AND REFRIGERANT MASS FLOW RATE OVER 24
Figure 30 depicts the hourly profile of temperature differential (TD) between the
saturated evaporating temperature and DAT. As depicted, the coil TD was highest at
the end of defrost periods, and started to decline as the refrigeration period began.
The average coil TD over 24-hours of testing remained around 4oF. The observed coil
TD profile is attributed to maintaining a fixed suction pressure, which resulted in
maintaining a constant evaporator temperature of 22oF, coupled with variations in
DAT.
10
9
Evap Coil TD (F)
8
7
6
5
4
3
2
1
0
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
FIGURE 30. HOURLY PROFILE OF EVAPORATOR COIL TEMPERATURE DIFFERENCE (TD) OVER 24 HOURS – HILL
PHOENIX DISPLAY CASE
The evaporator coil superheat and total system sub-cooling remained relatively
constant during the refrigeration periods (Figure 31). However, some variations were
observed when the system was approaching defrost and after defrost periods. The
Page 42
June 2009
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Evap Coil Superheat &
Total Subcooling (F)
average evaporator superheat remained around 5oF, and average total system
subcooling remained around 30oF.
32
30
28
26
24
22
20
18
16
14
12
10
8
6
4
2
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Evap Coil Superheat
Total Subcooling
FIGURE 31. HOURLY PROFILE OF EVAPORATOR COIL SUPERHEAT AND TOTAL SYSTEM SUB-COOLING OVER 24
The hourly profile for the display case’s evaporator fan motors and lighting system
power usage is shown in Figure 32. As shown, the evaporator fan motors and
lighting system power consumption remained unchanged over the entire test period.
140
120
Power (W)
100
80
60
40
20
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
17:12:39
16:10:39
15:08:39
14:06:39
13:04:39
12:02:39
11:00:39
9:58:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
2:44:39
0
Case Fans
Case Lighting
FIGURE 32. TWO-MINUTE PROFILE OF CASE LIGHTING AND EVAPORATOR FAN MOTOR POWER OVER 24 HOURS –
Figure 33 depicts the total system power and the total power usage by end-use over
the entire test period. The case’s evaporator fan motors power and lighting system
power were approximately 0.06 kW and 0.11 kW, respectively. The largest
contributor to the total system power was the refrigeration system compressor with
2.24 kW. The total power usage over the entire test period equaled 2.44 kW.
Page 43
June 2009
ET 06.07
3.0
2.44
Average Power (kW)
[over 24-hours]
2.5
2.24
2.0
1.5
1.0
0.5
0.06
0.11
Evap. Fan
Lighting
0.0
Compressor
Total System
End-Use
FIGURE 33. AVERAGE TOTAL AND END-USE POWER OVER 24 HOURS – HILL PHOENIX DISPLAY CASE
Left Front
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
18:14:39
Center Rear
17:12:39
16:10:39
14:06:39
13:04:39
12:02:39
Right Rear
11:00:39
9:58:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
Left Rear
Right Front
19:16:39
Center Front
15:08:39
45
43
41
39
37
35
33
31
29
27
25
2:44:39
Bottom Shelf (shelf #1)
Product Temp (F)
Additionally, the 2-minute profile of product temperatures at six locations inside the
display case for each shelf was monitored (Figure 34 through Figure 38). A review of
these figures reveals that there was a variation in product temperature profiles
depending on the product location, and it varied among shelves. However, the rear
products were lower in temperature than the front products, as expected. Also,
products located at the left side inside the case were always lower in temperature
than those located at the right and center locations.
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
FIGURE 34. TWO-MINUTE PROFILE OF PRODUCT TEMPERATURE AT SIX DIFFERENT LOCATIONS FOR BOTTOM
SHELF OVER 24 HOURS – HILL PHOENIX DISPLAY CASE
Page 44
June 2009
Left Front
Center Front
39
37
35
Right Front
33
31
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
Center Rear
17:12:39
16:10:39
14:06:39
13:04:39
12:02:39
11:00:39
Right Rear
9:58:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
Left Rear
15:08:39
29
27
25
2:44:39
Shelf #2
Product Temp (F)
45
43
41
ET 06.07
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
FIGURE 35. TWO-MINUTE PROFILE OF PRODUCT TEMPERATURE AT SIX DIFFERENT LOCATIONS FOR SECOND
Left Front
Center Front
39
37
35
Right Front
33
31
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
Center Rear
17:12:39
16:10:39
14:06:39
13:04:39
12:02:39
11:00:39
Right Rear
9:58:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
Left Rear
15:08:39
29
27
25
2:44:39
Shelf #3
Product Temp (F)
45
43
41
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
FIGURE 36. TWO-MINUTE PROFILE OF PRODUCT TEMPERATURE AT SIX DIFFERENT LOCATIONS FOR THIRD SHELF
Page 45
June 2009
45
39
37
Left Front
35
33
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
Center Rear
17:12:39
16:10:39
14:06:39
13:04:39
12:02:39
11:00:39
Right Rear
9:58:39
8:56:39
7:54:39
6:52:39
5:50:39
4:48:39
3:46:39
Left Rear
15:08:39
31
29
27
25
2:44:39
Shelf #4
Product Temp (F)
Right Front
Center Front
43
41
ET 06.07
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
FIGURE 37. TWO-MINUTE PROFILE OF PRODUCT TEMPERATURE AT SIX DIFFERENT LOCATIONS FOR FOURTH
43
Right Front
Left Front
Top Shelf (shelf #5)
Product Temp (F)
41
39
37
Center Front
35
33
31
29
2:30:39
1:28:39
0:26:39
23:24:39
22:22:39
21:20:39
20:18:39
19:16:39
18:14:39
Center Rear
17:12:39
16:10:39
15:08:39
14:06:39
13:04:39
12:02:39
11:00:39
Right Rear
9:58:39
8:56:39
6:52:39
5:50:39
4:48:39
3:46:39
2:44:39
7:54:39
Left Rear
27
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
FIGURE 38. TWO-MINUTE PROFILE OF PRODUCT TEMPERATURE AT SIX DIFFERENT LOCATIONS FOR TOP SHELF
Figure 39 depicts the average of all six product temperatures for each shelf. As
illustrated, the product temperatures were lowest for the top shelf and highest for
the fourth shelf (one shelf below the top shelf). Figure 39 also shows that the
average product temperatures had similar profiles regardless of variations in
temperature magnitudes. The variations in temperature magnitude are attributed to
defrost periods, and, in fact, the products experienced a temperature swing of 2oF to
3oF as a result of the defrost periods.
Page 46
June 2009
ET 06.07
Shelf #4
Shelf #3
4:36:39
7:24:39
Average Product
Temperatures (F)
36.5
Shelf #2
Bottom Shelf
Top Shelf
35.5
34.5
2:04:39
1:08:39
0:12:39
23:16:39
22:20:39
21:24:39
20:28:39
19:32:39
18:36:39
17:40:39
16:44:39
15:48:39
14:52:39
13:56:39
13:00:39
12:04:39
11:08:39
9:16:39
10:12:39
8:20:39
6:28:39
5:32:39
3:40:39
2:44:39
33.5
Top Shelf
Shelf #2
Shelf #3
Shelf #4
Bottom Shelf
FIGURE 39. AVERAGE PRODUCT TEMPERATURES FOR EACH SHELF OVER 24 HOURS – HILL PHOENIX DISPLAY
CASE
The coldest product temperature was 27.6oF, and the warmest was about 41.8oF
(Figure 40). Average coldest and warmest product temperatures were 28.7oF and
41.2oF, respectively. Averaging all of the product simulators yielded an average
product temperature of 34.9oF.
45
41.2
41.8
Warmest Test
Simulator Average
Temp.
Warmest Test
Simulator Temp.
Product Temp (F)
[including defrost periods]
40
34.9
35
28.7
30
27.6
25
20
15
10
5
0
Average Product
Temp. of All Test
Simulators
Coldest Test
Simulator Average
Temp.
Coldest Test
Simulator Temp.
FIGURE 40. AVERAGE, COLDEST AND WARMEST PRODUCT TEMPERATURES OVER 24 HOURS – HILL PHOENIX
DISPLAY CASE
Page 47
June 2009
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Additionally, the collected test data was compared to the manufacturer’s published
data. The results are summarized in Table 4. Although the manufacturer’s
recommended defrost frequency was four times per day, during preliminary testing it
was evident that due to heavy frost formation on the evaporator coil, the case
operation was hampered. Accordingly, the manufacturer recommended increasing
the defrost frequency from four to six times per day. As a result, the cooling load
was higher than manufacturer’s data.
TABLE 4.
COMPARATIVE SUMMARY OF TEST DATA AND MANUFACTURER’S PUBLISHED DATA – HILL PHOENIX
DISPLAY CASE
KEY PARAMETERS
MANUFACTURER DATA
TEST DATA (AVERAGE)
1,570
2,015
22
23
6–8
5
Discharge Air Temperature (oF)
30
30
Return Air Temperature (oF)
44
45
4
6
42
28
Cooling Load per Linear-feet
(Btu/hr/ft)
Saturated Evaporating
Temperature (oF)
Superheat Set Point (oF)
Defrost per Day
Defrost Duration (minutes)
As described earlier (see Figure 13), this Hill Phoenix display case model can be
manufactured with different optional front extensions. Again, the one that was tested
in this project was a standard model that had two-and-a-half inch front extension.
Therefore, the effects of the highest available optional front extension, which was
seven-and-a-half inch (see Figure 13), on the infiltration load for this display case
model was investigated.
The investigation involved using an artificial neural network (ANN) program that was
newly developed, and validated by numerical and experimental data [Ref .4]. This
analysis was conducted by collaborating with Kettering University. The results of this
analysis indicated that when the vertical distance between the discharge and return
air grille was reduced by 5-inch, from 2.5-inch to 7.5-inch, the infiltration was
estimated to decrease between 2.5% to 3%. Subsequently, a reduction in
compressor power can be expected.
HUSSMANN DISPLAY CASE (M5X-GEP)
The performance of the Hussmann display case was evaluated under ASHRAE 72-05
DB and RH was 75oF and 55.1%, respectively, which corresponded to a WB of 64oF.
Page 48
June 2009
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80
75
70
65
60
55
50
45
40
35
30
25
20
85
75
Room RH (%)
65
55
45
35
13:51:25
12:47:25
11:43:25
9:35:25
10:39:25
8:31:25
7:27:25
6:23:25
5:19:25
4:15:25
3:11:25
2:07:25
1:03:25
23:59:25
22:55:25
21:51:25
20:47:25
19:43:25
18:39:25
17:35:25
16:31:25
15:27:25
25
15
5
14:23:25
Room Temp (F)
Room Temp
Room RH
OVER 24 HOURS – HUSSMANN DISPLAY CASE
66
64
220
62
60
210
200
58
56
54
190
180
52
50
14:09:25
13:07:25
12:05:25
11:03:25
10:01:25
8:59:25
7:57:25
6:55:25
5:53:25
4:51:25
3:49:25
2:47:25
1:45:25
0:43:25
23:41:25
22:39:25
21:37:25
20:35:25
19:33:25
18:31:25
17:29:25
16:27:25
15:25:25
170
230
14:23:25
In order to maintain a DAT of 30oF or an evaporator temperature of 26oF, as
specified by the manufacturer, the test rack controller was programmed to run at a
fixed suction pressure of 61 psig (Figure 42). The rack controller was also
programmed to run at a fixed discharge pressure of 220 psig or 95oF SCT. Figure 42
illustrates the two-minute profile of suction and discharge pressures over the entire
test period.
Discharge Pressure
Suction Pressure
FIGURE 42. TWO-MINUTE PROFILE OF SUCTION AND DISCHARGE PRESSURES OVER 24 HOURS – HUSSMANN
DISPLAY CASE
temperature. As shown, the refrigeration system maintained a relatively constant
average DAT of 30oF during the entire test period. In addition, the average return air
temperature was 38.5oF.
Page 49
June 2009
ET 06.07
57
Average Discharge &
Return Air Temp (F)
53
49
45
41
37
33
29
14:09:25
13:07:25
12:05:25
11:03:25
10:01:25
8:59:25
7:57:25
6:55:25
5:53:25
4:51:25
3:49:25
2:47:25
1:45:25
0:43:25
23:41:25
22:39:25
21:37:25
20:35:25
19:33:25
18:31:25
17:29:25
16:27:25
15:25:25
14:23:25
25
Discharge Air Temp
Return Air Temp
HUSSMANN DISPLAY CASE
14:09:25
13:07:25
12:05:25
11:03:25
10:01:25
8:59:25
7:57:25
6:55:25
5:53:25
4:51:25
3:49:25
2:47:25
1:45:25
0:43:25
23:41:25
22:39:25
21:37:25
20:35:25
19:33:25
18:31:25
17:29:25
16:27:25
15:25:25
100
90
80
70
60
50
40
30
20
10
0
14:23:25
Mass of Collected
Condensate (lbs)
during off-cycle defrost periods. The stepped (horizontal) profiles indicate the
moisture collected during the refrigeration run time between each of the four defrost
periods. The vertical (or sloped) profiles, on the other hand, indicate the melted ice
during each defrost period.
FIGURE 44. TWO-MINUTE PROFILE OF COLLECTED CONDENSATE OVER 24 HOURS – HUSSMANN DISPLAY CASE
Page 50
June 2009
ET 06.07
Over 24-hours (lbs)
100
90
75.95
80
70
60
46.37
50
40
24.72
30
20
4.81
10
0
during defrosts
Mass of water vapor
condensed from air during
defrosts
Mass of water vapor
periods
Total measured
condensate
FIGURE 45. BREAKDOWN OF CONDENSATE COLLECTED OVER 24 HOURS – HUSSMANN DISPLAY CASE
85
75
65
55
45
35
25
Inside Case RH (%)
95
80
75
70
65
60
55
50
45
40
35
30
25
20
13:51:25
12:47:25
11:43:25
9:35:25
10:39:25
8:31:25
7:27:25
6:23:25
5:19:25
4:15:25
3:11:25
2:07:25
1:03:25
23:59:25
22:55:25
21:51:25
20:47:25
19:43:25
18:39:25
17:35:25
16:31:25
15:27:25
15
14:23:25
temperature and humidity inside the case were lowered.
Inside Case Temp
Inside Case RH
warm and humid air into the case during defrost periods. The lowest cooling load
was observed prior to initiating a defrost period. The average cooling load was 1,578
Btu/hr/ft.
Page 51
June 2009
ET 06.07
2,500
2,400
2,300
2,200
2,100
2,000
1,900
1,800
1,700
1,600
1,500
1,400
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
HOURS – HUSSMANN DISPLAY CASE
the conduction, 551 Btu/hr, corresponding to 4% of the total cooling load. The
internal load that consisted of the display case’s lighting system and the evaporator
fan motors accounted for roughly 6% and radiation for about 8% of the total cooling
load.
16,000
14,000
12,000
10,339
10,000
8,000
6,000
4,000
2,000
551
1,001
Conduction
Radiation
730
0
Infiltration
fans)
FIGURE 48. COOLING LOAD BY COMPONENT OVER 24 HOURS – HUSSMANN DISPLAY CASE
Page 52
June 2009
fans)
5.8%
Conduction
4.4%
ET 06.07
Radiation
8.0%
Infiltration
81.9%
FIGURE 49. PERCENTAGE BREAKDOWN OF THE COOLING LOAD COMPONENTS OVER 24 HOURS – HUSSMANN
DISPLAY CASE
18,000
16,000
14,000
12,621
12,622
11,371
running cycle)
12,000
10,000
8,000
6,000
4,000
2,000
0
CYCLE – HUSSMANN DISPLAY CASE
The mass flow rate of refrigerant was observed to decline slightly during each
running cycle (Figure 51). The flow rate was highest at the end of defrost and lowest
prior to initiation of defrost. This observed profile in refrigerant mass flow rate was
attributed to the change in total cooling load of the case coupled with maintaining a
fixed suction pressure during the entire test period.
Page 53
June 2009
ET 06.07
7
Refrigerant Mass
Flow Rate (lb/min)
6
5
4
3
2
y = -0.0003x + 3.6836
R2 = 0.0127
1
13:05:25
12:15:25
10:35:25
11:25:25
8:55:25
9:45:25
7:27:25
5:47:25
6:37:25
4:07:25
4:57:25
3:17:25
0:59:25
2:27:25
23:19:25
0:09:25
22:29:25
20:49:25
21:39:25
19:23:25
17:43:25
18:33:25
16:03:25
16:53:25
15:13:25
14:23:25
0
FIGURE 51. TWO-MINUTE PROFILE OF REFRIGERANT MASS FLOW RATE OVER 24 HOURS – HUSSMANN DISPLAY
CASE
flow rate.
Refrigerant Mass
Flow Rate (lb/min)
y = -0.0003x + 3.6836
R2 = 0.0127
6
5
5
4
4
3
3
2
2
y = -1E-04x + 1.9593
R2 = 0.0282
1
1
0
10:35:25
11:25:25
12:15:25
13:05:25
4:07:25
4:57:25
5:47:25
6:37:25
7:27:25
8:55:25
9:45:25
21:39:25
22:29:25
23:19:25
0:09:25
0:59:25
2:27:25
3:17:25
15:13:25
16:03:25
16:53:25
17:43:25
18:33:25
19:23:25
20:49:25
14:23:25
0
6
7
Compressor Power
FIGURE 52. TWO-MINUTE PROFILE OF COMPRESSOR POWER AND REFRIGERANT MASS FLOW RATE OVER 24
DAT.
Page 54
June 2009
ET 06.07
10
Evap Coil TD (F)
8
6
4
2
0
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
FIGURE 53. HOURLY PROFILE OF EVAPORATOR COIL TEMPERATURE DIFFERENCE (TD) OVER 24 HOURS –
The evaporator coil superheat and total system subcooling remained relatively
average evaporator superheat remained around 6oF, and average total system
34
32
30
28
26
24
22
20
18
16
14
12
10
8
6
4
2
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Evap Coil Superheat
Total Subcooling
FIGURE 54. HOURLY PROFILE OF EVAPORATOR COIL SUPERHEAT AND TOTAL SYSTEM SUBCOOLING OVER 24
The hourly profile for the display case’s evaporator fan motors and lighting system
power usage is shown in Figure 55. As shown, the lighting system power
consumption remained unchanged over the entire test period. The evaporator fan
Page 55
June 2009
ET 06.07
motors power consumption, however, was increased prior to initiation of defrost due
to frost build up on the evaporator coils.
140
120
Power (W)
100
80
60
40
20
14:09:25
13:07:25
12:05:25
11:03:25
10:01:25
8:59:25
7:57:25
6:55:25
5:53:25
4:51:25
3:49:25
2:47:25
1:45:25
0:43:25
23:41:25
22:39:25
21:37:25
20:35:25
19:33:25
18:31:25
17:29:25
16:27:25
15:25:25
14:23:25
0
Case Fans
Case Lighting
FIGURE 55. TWO-MINUTE PROFILE OF CASE LIGHTING AND EVAPORATOR FAN MOTOR POWER OVER 24 HOURS –
power were approximately 0.09 kW and 0.12 kW, respectively. The largest
contributor to the total system power was the refrigeration system compressor with
1.93 kW. The total power usage over the entire test period equaled 2.15 kW.
2.5
2.15
1.93
Average Power (kW)
[over 24-hours]
2.0
1.5
1.0
0.5
0.09
0.12
0.0
Evap. Fan
Lighting
Compressor
Total System
End-Use
FIGURE 56. AVERAGE TOTAL AND END-USE POWER OVER 24 HOURS – HUSSMANN DISPLAY CASE
Figure 57 through Figure 60 revealed that there was a variation in product
temperature profiles depending on the product location, and it varied among shelves.
However, the rear products were lower in temperature than the front products, as
expected.
Page 56
June 2009
43
Product Temp (F)
Left Front
Center Front
41
ET 06.07
Right Front
39
37
35
33
31
Left Rear
29
Center Rear
Right Rear
27
14:09:25
13:07:25
12:05:25
11:03:25
10:01:25
8:59:25
7:57:25
6:55:25
5:53:25
4:51:25
3:49:25
2:47:25
1:45:25
0:43:25
23:41:25
22:39:25
21:37:25
20:35:25
19:33:25
18:31:25
17:29:25
16:27:25
15:25:25
14:23:25
25
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
SHELF OVER 24 HOURS – HUSSMANN DISPLAY CASE
43
Center Front
41
Left Front
Right Front
Shelf #2
Product Temp (F)
39
37
35
33
31
29
Right Rear
Center Rear
27
Left Rear
14:09:25
13:07:25
12:05:25
11:03:25
10:01:25
8:59:25
7:57:25
6:55:25
5:53:25
4:51:25
3:49:25
2:47:25
1:45:25
0:43:25
23:41:25
22:39:25
21:37:25
20:35:25
19:33:25
18:31:25
17:29:25
16:27:25
15:25:25
14:23:25
25
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
SHELF OVER 24 HOURS – HUSSMANN DISPLAY CASE
Page 57
June 2009
43
Shelf #3
Product Temp (F)
Left Front
Center Front
41
ET 06.07
Right Front
39
37
35
33
31
29
Center Rear
Left Rear
Right Rear
27
14:09:25
13:07:25
12:05:25
11:03:25
10:01:25
8:59:25
7:57:25
6:55:25
5:53:25
4:51:25
3:49:25
2:47:25
1:45:25
0:43:25
23:41:25
22:39:25
21:37:25
20:35:25
19:33:25
18:31:25
17:29:25
16:27:25
15:25:25
14:23:25
25
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
43
Center Front
Top Shelf (shelf #4)
Product Temp (F)
41
Right Front
Left Front
39
37
35
33
31
29
Left Rear
Right Rear
27
Center Rear
14:09:25
13:07:25
12:05:25
11:03:25
10:01:25
8:59:25
7:57:25
6:55:25
5:53:25
4:51:25
3:49:25
2:47:25
1:45:25
0:43:25
23:41:25
22:39:25
21:37:25
20:35:25
19:33:25
18:31:25
17:29:25
16:27:25
15:25:25
14:23:25
25
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
illustrated, the product temperatures were lowest for the third shelf (one shelf below
the top shelf) and highest for the bottom shelf. Figure 61 also shows that the
Page 58
June 2009
39.5
Top Shelf
Shelf #2
Bottom Shelf
ET 06.07
Shelf #3
Average Product
Temperatures (F)
38.5
37.5
36.5
35.5
34.5
33.5
13:43:25
12:47:25
11:51:25
10:55:25
9:59:25
9:03:25
8:07:25
7:11:25
6:15:25
5:19:25
4:23:25
3:27:25
2:31:25
1:35:25
0:39:25
23:43:25
22:47:25
21:51:25
20:55:25
19:59:25
19:03:25
18:07:25
17:11:25
16:15:25
15:19:25
14:23:25
32.5
Top Shelf
Shelf #2
Shelf #3
Bottom Shelf
FIGURE 61. AVERAGE PRODUCT TEMPERATURES FOR EACH SHELF OVER 24 HOURS – HUSSMANN DISPLAY CASE
The coldest product temperature was 29.3oF, and the warmest was about 39.8oF
(Figure 62). Average coldest and warmest product temperatures were 30.9oF and
38.3oF, respectively. Averaging all of the product simulators yielded an average
product temperature of 35.2oF.
45
40
38.3
39.8
Warmest Test
Simulator Average
Temp.
Warmest Test
Simulator Temp.
Product Temp (F)
35.2
35
30.9
30
29.3
25
20
15
10
5
0
Average Product
Temp. of All Test
Simulators
Coldest Test
Simulator Average
Temp.
Coldest Test
Simulator Temp.
FIGURE 62. AVERAGE, COLDEST AND WARMEST PRODUCT TEMPERATURES OVER 24 HOURS – HUSSMANN
DISPLAY CASE
data. The results are summarized in Table 5. Overall, the obtained test results were
in close agreement with the manufacturer’s published data.
Page 59
June 2009
TABLE 5.
ET 06.07
COMPARATIVE SUMMARY OF TEST DATA AND MANUFACTURER’S PUBLISHED DATA – HUSSMANN
DISPLAY CASE
KEY PARAMETERS
MANUFACTURER DATA
TEST DATA (AVERAGE)
1,370
1,578
Temperature (oF)
26
26
30
30
4
4
35
35
(lb/ft/day)
9.0
9.5
(Btu/hr/ft)
Defrost per Day
Since it was possible to remove the glass-front extension, additional test run was
initiated to better understand and quantify the benefits of the glass-front extension.
The display case was tested without the glass-front extension for a 24-hour period
while maintaining the test chamber at 75oF DB and 55% RH during the entire test
period. Additionally, the refrigeration system was set to provide a DAT of 30oF as
specified by the manufacturer while keeping an SCT of 95oF.
The results indicated that the infiltration load increased from 10,359 Btu/hr to
13,326 Btu/hr, or by 2,967 Btu/hr, while other cooling load components remained
fairly unchanged. Accordingly, the compressor power demand increased from 1.93
kW to 2.27 kW, or by 0.33 kW. In other words, removing the 7.5-inch glass-front
extension resulted in a 28% increase in infiltration load and an 18% increase in
compressor power demand. This finding also verified and supported a previous study
that revealed reducing the vertical distance between the discharge and the return air
grille reduces the infiltration rate or load of the display case [Ref. 4].
TYLER DISPLAY CASE (N6DHPACLA)
The performance of the Tyler display case was evaluated under ASHRAE 72-05
DB and RH was 75oF and 54.8%, respectively, which corresponded to a WB of
63.9oF.
Page 60
June 2009
ET 06.07
80
75
70
65
60
55
50
45
40
35
30
25
20
85
65
55
45
35
25
Room RH (%)
75
15
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
10:30:38
9:28:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
3:16:38
2:14:38
1:12:38
0:10:38
23:08:38
22:06:38
21:04:38
5
20:02:38
Room Temp (F)
Room Temp
Room RH
OVER 24 HOURS – TYLER DISPLAY CASE
230
64
220
62
210
60
200
58
56
190
54
180
52
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
10:30:38
9:28:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
3:16:38
2:14:38
1:12:38
0:10:38
23:08:38
22:06:38
50
21:04:38
170
66
20:02:38
In order to maintain a DAT of 34.5oF as specified by the manufacturer, the test rack
controller was programmed to run at a fixed suction pressure of 59 psig (Figure 64).
The rack controller was also programmed to run at a fixed discharge pressure of 220
psig or 95oF SCT. Figure 64 illustrates the two-minute profile of suction and
discharge pressures over the entire test period.
Discharge Pressure
Suction Pressure
FIGURE 64. TWO-MINUTE PROFILE OF SUCTION AND DISCHARGE PRESSURES OVER 24 HOURS – TYLER DISPLAY
CASE
temperature (RAT). It can be noted that as the refrigeration period continued, the
DAT as well as the RAT started to increase until the next defrost period was initiated.
Nonetheless, the average DAT was around 35.5oF, which was 1oF above the
manufacturer’s specified DAT. In addition, the average RAT was 51.5oF.
Figure 66 shows that the average DAT at five different locations along the discharge
air grille varied from 34oF to 37oF. Therefore, maintaining an average DAT of 34.5oF
Page 61
June 2009
ET 06.07
could not be achieved. However, an average DAT of 34.8oF was achieved at the
center location, which was very close to the manufacturer’s specified DAT of 34.5oF.
60
Average Discharge &
Return Air Temp (F)
55
50
45
40
35
30
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
10:30:38
9:28:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
3:16:38
2:14:38
1:12:38
0:10:38
23:08:38
22:06:38
21:04:38
20:02:38
25
Discharge Air Temp
Return Air Temp
TYLER DISPLAY CASE
Discharge Air Temperatures (F)
38
37
37.0
36.3
36
35.5
35.5
34.8
35
33.8
34
33
32
Left Location
Left Center
Location
Center Location
Right Center
Location
Right Location
Average of all
Five Locations
FIGURE 66. INDIVIDUAL AND AVERAGE DISCHARGE AIR TEMPERATURE OVER 24 HOURS – TYLER DISPLAY CASE
during off-cycle defrost periods. The sloped horizontal profiles indicate the moisture
collected during the refrigeration run time between each of the six defrost periods.
The sloped vertical profiles, on the other hand, indicate the melted ice during each
defrost period.
Page 62
June 2009
ET 06.07
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
9:28:38
10:30:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
3:16:38
2:14:38
1:12:38
0:10:38
23:08:38
22:06:38
21:04:38
100
90
80
70
60
50
40
30
20
10
0
20:02:38
Mass of Collected
Condensate (lbs)
FIGURE 67. TWO-MINUTE PROFILE OF COLLECTED CONDENSATE OVER 24 HOURS – TYLER DISPLAY CASE
Over 24-hours (lbs)
100
90.36
90
80
70
60
54.51
50
40
32.33
30
20
10
3.52
0
during defrosts
Mass of water vapor
condensed from air
during defrosts
Mass of water vapor
periods
Total measured
condensate
FIGURE 68. BREAKDOWN OF CONDENSATE COLLECTED OVER 24 HOURS – TYLER DISPLAY CASE
temperature and humidity inside the case were lowered. However, as the
refrigeration period continued, the temperature and humidity levels started to
Page 63
June 2009
ET 06.07
Inside Case RH (%)
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
9:28:38
10:30:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
3:16:38
2:14:38
1:12:38
0:10:38
23:08:38
22:06:38
120
110
100
90
80
70
60
50
40
30
20
10
0
21:04:38
80
75
70
65
60
55
50
45
40
35
30
25
20
20:02:38
increase until the next defrost period was initiated. This observation is attributed to a
decrease in evaporator coil capacity over time as a result of frost formation on the
coil.
Inside Case Temp
Inside Case RH
TYLER DISPLAY CASE
warm and humid air into the case during defrost periods. The lowest cooling load
was observed prior to initiating defrost periods. The average cooling load was 1,708
Btu/hr/ft.
2,500
2,400
2,300
2,200
2,100
2,000
1,900
1,800
1,700
1,600
1,500
1,400
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
HOURS – TYLER DISPLAY CASE
Page 64
June 2009
ET 06.07
the display case’s lighting system and evaporator fan motors (internal load), 476
Btu/hr, corresponding to about 4% of the total cooling load. The conduction
accounted for roughly 4% and radiation for about 7% of the total cooling load.
16,000
14,000
11,716
12,000
10,000
8,000
6,000
4,000
2,000
496
973
Conduction
Radiation
476
0
Infiltration
fans)
FIGURE 71. COOLING LOAD BY COMPONENT OVER 24 HOURS – TYLER DISPLAY CASE
fans)
3.5%
Conduction
3.6%
Radiation
7.1%
Infiltration
85.8%
FIGURE 72. PERCENTAGE BREAKDOWN OF THE COOLING LOAD COMPONENTS OVER 24 HOURS – TYLER DISPLAY
CASE
Page 65
June 2009
ET 06.07
18,000
16,000
13,662
13,662
14,000
12,596
12,000
10,000
8,000
6,000
4,000
2,000
0
running cycle)
CYCLE – TYLER DISPLAY CASE
The mass flow rate of refrigerant was observed to decline during each running cycle
(Figure 74). The flow rate was highest at the end of defrost and lowest prior to
initiation of defrost periods. This observed profile in refrigerant mass flow rate was
attributed to the change in total cooling load of the case coupled with maintaining a
fixed suction pressure during the entire test period.
8
Refrigerant Mass
Flow Rate (lb/min)
7
6
5
4
3
y = -0.0003x + 3.8984
R2 = 0.0126
2
1
19:40:38
18:50:38
18:00:38
17:10:38
16:20:38
15:08:38
14:18:38
13:28:38
12:38:38
11:26:38
9:46:38
10:36:38
8:56:38
8:06:38
6:52:38
6:02:38
5:12:38
4:22:38
3:08:38
2:18:38
1:28:38
0:38:38
23:24:38
22:34:38
21:44:38
20:54:38
20:04:38
0
FIGURE 74. TWO-MINUTE PROFILE OF REFRIGERANT MASS FLOW RATE OVER 24 HOURS – TYLER DISPLAY CASE
flow rate.
Page 66
June 2009
ET 06.07
7
y = -0.0003x + 3.8984
R2 = 0.0126
6
6
5
5
4
4
3
3
2
2
y = 4E-05x + 2.0698
R2 = 0.0035
19:40:38
17:10:38
18:00:38
18:50:38
14:18:38
15:08:38
16:20:38
11:26:38
12:38:38
13:28:38
8:06:38
8:56:38
9:46:38
10:36:38
0
5:12:38
6:02:38
6:52:38
20:54:38
21:44:38
22:34:38
20:04:38
0
1
2:18:38
3:08:38
4:22:38
1
23:24:38
0:38:38
1:28:38
Refrigerant Mass
Flow Rate (lb/min)
7
Compressor Power
FIGURE 75. TWO-MINUTE PROFILES OF COMPRESSOR POWER AND REFRIGERANT MASS FLOW RATE OVER 24
DAT.
10
9
Evap Coil TD (F)
8
7
6
5
4
3
2
1
0
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
FIGURE 76. HOURLY PROFILE OF EVAPORATOR COIL TEMPERATURE DIFFERENCE (TD) OVER 24 HOURS – TYLER
DISPLAY CASE
The evaporator coil superheat and total system subcooling remained relatively
Page 67
June 2009
ET 06.07
average evaporator superheat remained around 14oF, and the average total system
70
65
60
55
50
45
40
35
30
25
20
15
10
5
0
1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Evap Coil Superheat
Total Subcooling
FIGURE 77. HOURLY PROFILE OF EVAPORATOR COIL SUPERHEAT AND TOTAL SYSTEM SUBCOOLING OVER 24
The hourly profile for the display case’s evaporator fan motors, secondary or ambient
fan motors, and lighting system power usage is shown in Figure 78. As shown, the
evaporator fan motors, secondary or ambient fan motors, and lighting system power
consumption remained unchanged over the entire test period.
120
100
Case Lighting
Power (W)
80
60
Evaporator Fans
40
20
Ambient Fans
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
10:30:38
9:28:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
3:16:38
2:14:38
1:12:38
0:10:38
23:08:38
22:06:38
21:04:38
20:02:38
0
Evaporator Fans
Case Lighting
Ambient Fans
FIGURE 78. HOURLY PROFILE OF CASE LIGHTING AND EVAPORATOR FAN MOTOR POWER OVER 24 HOURS –
TYLER DISPLAY CASE
power were approximately 0.03 kW and 0.11 kW, respectively. The secondary or
ambient fan motors power was about 0.02 kW. The largest contributor to the total
Page 68
June 2009
ET 06.07
system power was the refrigeration system compressor with 2.12 kW. The total
power usage over the entire test period equaled 2.28 kW.
2.5
2.28
2.12
Average Power (kW)
[over 24-hours]
2.0
1.5
1.0
0.5
0.03
0.02
0.11
Evap. Fans
Ambient Fans
Lighting
0.0
Compressor
Total System
End-Use
FIGURE 79. AVERAGE TOTAL AND END-USE POWER OVER 24 HOURS – TYLER DISPLAY CASE
50
48
46
44
42
40
38
36
34
32
30
Center Front
Left Front
Right Front
Center Rear
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
10:30:38
9:28:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
3:16:38
2:14:38
1:12:38
Right Rear
0:10:38
23:08:38
22:06:38
21:04:38
Left Rear
20:02:38
Product Temp (F)
these figures revealed that there was a variation in product temperature profiles
depending on the product location, and it varied among shelves. However, the rear
products were lower in temperature than the front products, as expected.
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
SHELF OVER 24 HOURS – TYLER DISPLAY CASE
Page 69
June 2009
46
Left Front
Center Front
ET 06.07
Right Front
42
40
38
36
34
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
9:28:38
8:26:38
Right Rear
7:24:38
6:22:38
3:16:38
2:14:38
1:12:38
0:10:38
23:08:38
22:06:38
21:04:38
20:02:38
5:20:38
Center Rear
Left Rear
30
10:30:38
32
4:18:38
Shelf #2
Product Temp (F)
44
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
46
Right Front
Center Front
Shelf #3
Product Temp (F)
44
42
40
Left Front
38
36
34
32
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
Center Rear
10:30:38
9:28:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
3:16:38
Right Rear
2:14:38
1:12:38
23:08:38
22:06:38
21:04:38
20:02:38
0:10:38
Left Rear
30
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
Page 70
June 2009
44
ET 06.07
Left Front
Center Front
Shelf #4
Product Temp (F)
42
40
Right Front
38
36
34
32
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
Center Rear
10:30:38
9:28:38
8:26:38
7:24:38
6:22:38
5:20:38
4:18:38
Right Rear
3:16:38
2:14:38
0:10:38
23:08:38
22:06:38
21:04:38
20:02:38
1:12:38
Left Rear
30
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
FIGURE 83. TWO-MINUTE PROFILE OF PRODUCT TEMPERATURE AT SIX DIFFERENT LOCATIONS FOR FOURTH
50
Right Front
Left Front
44
42
40
38
Center Front
36
34
19:48:38
18:46:38
17:44:38
16:42:38
15:40:38
14:38:38
13:36:38
12:34:38
11:32:38
Center Rear
10:30:38
9:28:38
7:24:38
6:22:38
5:20:38
4:18:38
Right Rear
3:16:38
2:14:38
0:10:38
23:08:38
22:06:38
21:04:38
20:02:38
Left Rear
8:26:38
32
30
1:12:38
Top Shelf
Product Temp (F)
48
46
Left Rear
Left Front
Center Rear
Center Front
Right Rear
Right Front
illustrated, the product temperatures were lowest for the fourth shelf (one shelf
below the top shelf) and highest for the bottom shelf. Figure 85 also shows that the
Page 71
June 2009
ET 06.07
44
Average Product
Temperatures (F)
Top Shelf
Bottom Shelf
42
40
Shelf #2
38
Shelf #3
Shelf #4
19:22:38
18:26:38
17:30:38
16:34:38
15:38:38
14:42:38
13:46:38
12:50:38
11:54:38
10:58:38
9:06:38
10:02:38
8:10:38
7:14:38
6:18:38
5:22:38
4:26:38
3:30:38
2:34:38
1:38:38
0:42:38
23:46:38
22:50:38
21:54:38
20:58:38
20:02:38
36
Bottom Shelf
Shelf #2
Shelf #3
Shelf #4
Top Shelf
FIGURE 85. AVERAGE PRODUCT TEMPERATURES FOR EACH SHELF OVER 24 HOURS – TYLER DISPLAY CASE
The coldest product temperature was 33.2oF, and the warmest was 48.4oF (Figure
86). Average coldest and warmest product temperatures were 33.8oF and 46.6oF,
respectively. Averaging all of the product simulators yielded an average product
temperature of 39.0oF.
60
48.4
Product Temp (F)
50
46.6
39.0
40
33.8
33.2
Coldest Test
Simulator Average
Temp.
Coldest Test
Simulator Temp.
30
20
10
0
Average Product
Temp. of All Test
Simulators
Warmest Test
Simulator Average
Temp.
Warmest Test
Simulator Temp.
FIGURE 86. AVERAGE, COLDEST AND WARMEST PRODUCT TEMPERATURES OVER 24 HOURS – TYLER DISPLAY
CASE
data. The results are summarized in Table 6. The obtained cooling load was about
61% higher than that specified by the manufacturer.
Page 72
June 2009
TABLE 6.
ET 06.07
COMPARATIVE SUMMARY OF TEST DATA AND MANUFACTURER’S PUBLISHED DATA – TYLER DISPLAY
CASE
KEY PARAMETERS
MANUFACTURER DATA
TEST DATA (AVERAGE)
1,059
1,708
28
23
34.5
35.5
6
6
18
18
(Btu/hr/ft)
Temperature (oF)
Defrost per Day
Additionally, in an attempt to lower the warmest product temperature to a desirable
level (equal to or below 41oF), additional test runs were initiated. These test runs
were initiated while maintaining the controlled environment room at 75oF DB and
55% RH. The test runs involved lowering the suction pressure by increments of 2 to
3 psig while monitoring the warmest product temperature inside the display case.
However, after lowering the suction pressure by about 10 psig, the warmest product
temperature was still higher than 41oF. In fact, lowering the suction pressure from
roughly 60 psig to 50 psig increased the warmest product temperature from 48oF to
55oF.
COMPARISON OF RESULTS
This section compares results obtained from testing all three standard high efficiency display
cases. These display cases were tested according to manufacturers’ specified DAT. The
comparison will establish and quantify key performance components such as cooling load,
product temperature, and compressor power and energy requirements for each of the three
tested display cases.
Figure 87 illustrates that the test chamber maintained relatively non-varying DB and RH
during the entire test period for all three tested display cases. As shown, the DB and RH
remained around 75oF and 55%, respectively, for all three tests.
Page 73
June 2009
Room Dry Bulb Temp (F) &
Relative Humidity (%)
90
85
80
75
70
65
60
55
50
45
40
35
30
25
20
ET 06.07
Avg Room Temp = 75.0 F
Avg Room RH = 55.3%
Avg Room RH = 55.1%
Avg Room RH = 54.8%
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Test Scenarios
(tested display cases, including defrost)
FIGURE 87.
COMPARISON OF TWO-MINUTE PROFILES OF THE CONTROLLED ENVIRONMENT ROOM DRY BULB AND RELATIVE
HUMIDITY OVER 24 HOURS – ALL THREE TEST SCENARIOS
Figure 88 depicts the discharge and suction pressures for all three tested display cases. The
refrigeration system maintained a fixed discharge pressure of about 220 psig or 95oF SCT
during all three test runs. The suction pressures were set to provide the case
manufacturers’ specified DAT. As shown in Figure 88, the refrigeration system maintained
fixed suction pressures during the entire test periods. In Figure 88, the SET and DAT are
also shown that correspond to the suction pressures. Although both Hill Phoenix and
Hussmann specified a DAT of 30oF for their cases, the Hussmann display case provided a
DAT of 30oF at 3 psig higher suction pressure than the Hill Phoenix display case.
140
260
Avg Discharge = 218.6 psig Avg Discharge = 217.9 psig Avg Discharge = 218.2 psig
(SCT = 94.5 F)
(SCT = 94.9 F)
(SCT = 95.0 F)
120
220
100
200
80
180
160
60
140
Avg Suction = 57.7 psig
120
Discharge Pressure (psig)
240
40
(SET = 23.0 F, DAT = 30.1 F) (SET = 26.5 F, DAT = 29.9 F) (SET = 23.3 F, DAT = 35.5 F)
100
20
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Test Scenarios
(tested display cases, excluding defrost)
FIGURE 88.
COMPARISON OF TWO-MINUTE PROFILES OF SUCTION AND DISCHARGE PRESSURES OVER 24 HOURS – ALL
THREE TEST SCENARIOS
Page 74
June 2009
ET 06.07
The average DAT and return air temperature (RAT) for all three tested display cases are
shown in Figure 89. The average DATs were maintained according to manufacturers’
specifications, expect for some minor deviations in the Tyler case. For the Tyler case, an
average DAT of 35.5oF was achieved, which was only 1oF higher than the specifications. The
difference between DAT and RAT was more significant for Hill Phoenix and Tyler display
cases. Specifically, both the Hill Phoenix and Tyler case experienced a 15oF to 16oF increase
in RAT. However, the Hussmann display case experienced only a 9oF increase in RAT.
90
70
Avg RAT = 38.5 F
Avg RAT = 51.1 F
80
65
60
70
55
50
60
45
50
40
35
40
30
30
20
Average Return Air Temp (F)
Average Discharge Air Temp (F)
Avg RAT = 44.6 F
25
Avg DAT = 30.1 F
Avg DAT = 29.9 F
Avg DAT = 35.5 F
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
20
Test Scenarios
FIGURE 89.
COMPARISON OF TWO-MINUTE PROFILES OF AVERAGE DISCHARGE AND RETURN AIR TEMPERATURES OVER 24
HOURS – ALL THREE TEST SCENARIOS
The average DAT and product temperature for all three tested display cases are shown in
Figure 90. The average DAT and product temperature for both Hill Phoenix and Hussmann
display cases was 30oF and 35oF, respectively. Although both display cases maintained an
average product temperature of about 35oF, product temperature swing was slightly higher
for the Hussmann case than the Hill Phoenix case. For the Tyler display case, the average
DAT was 35oF and average product temperature was 39oF.
Page 75
June 2009
Avg Prod Temp = 34.9 F
80
46
44
42
40
70
Temperature Sw ing
60
38
36
34
40
32
30
28
30
26
24
50
20
Avg DAT = 30.1 F
Avg DAT = 29.9 F
Avg DAT = 35.5 F
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Average Product Temp (F)
Average Discharge Air Temp (F)
90
ET 06.07
22
20
Test Scenarios
FIGURE 90.
COMPARISON OF TWO-MINUTE PROFILES OF AVERAGE DISCHARGE AIR TEMPERATURE AND PRODUCT
TEMPERATURE OVER 24 HOURS – ALL THREE TEST SCENARIOS
Figure 91 illustrates the coldest and warmest product temperature for all three tested
display cases over a 24-hour period. The coldest product temperature for all three display
cases was between 27oF and 34oF. The warmest product temperature for the Hussmann
case was about 40oF, which was 1oF lower than the Food and Drug Administration’s (FDA)
food code requirement of 41oF. The warmest product temperature for both Hill Phoenix and
Tyler cases, however, was above the FDAs food code requirement of 41oF. This difference
was more significant for the Tyler case, 7oF difference, than for the Hill Phoenix case, less
than 1oF difference. In other words, although the average product temperatures for both Hill
Phoenix and Tyler display cases were below 40oF (see Figure 90), the warmest product
temperatures were above 41oF (Figure 91).
Page 76
June 2009
ET 06.07
60
Coldest and Warmest
Product Temperature (F)
Coldest Product Temp
50
Warmest Product Temp
40
48.4
41.8
39.8
30
33.2
29.3
27.6
20
10
0
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Test Scnearios
FIGURE 91.
COMPARISON OF COLDEST AND WARMEST PRODUCT TEMPERATURES OVER 24 HOURS – ALL THREE TEST
SCENARIOS
70
65
Avg Refrig Effect = 57.6
Btu/lb
Btu/lb
Btu/lb
Avg Flow Rate = 4.7 lb/min
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Refrigeration
Effect (Btu/lb)
60
55
50
45
40
35
30
15
14
13
12
11
10
9
8
7
6
5
4
3
2
1
0
Refrigerant Mass
Flow Rate (lb/min)
Figure 92 depicts the refrigeration effect and refrigerant mass flow rate for all three tested
display cases. As described in the Data Analysis section, the refrigeration effect represents
the cooling capacity of the evaporator per pound of refrigerant flow. As shown in Figure 92,
despite slight variations, the refrigerant effect for all three display cases was around 58
Btu/lb. The lowest refrigerant mass flow rate (3.6 lb/min), however, was observed for the
Hussmann display case. This was mainly due to low cooling load requirements for this case.
The refrigerant mass flow rate of Hussmann was 23% (3.6 lb/min vs. 4.7 lb/min) lower than
the Hill Phoenix case and 8% (3.6 lb/min vs. 3.9 lb/min) lower than the Tyler case.
Test Scenarios
FIGURE 92.
COMPARISON OF TWO-MINUTE PROFILES OF REFRIGERATION EFFECT AND REFRIGERANT MASS FLOW RATE
OVER 24 HOURS – ALL THREE TEST SCENARIOS
Page 77
June 2009
ET 06.07
As discussed earlier, the compressor power use was entirely dependent on the refrigerant
mass flow rate variations due to maintaining a fixed suction and discharge pressures. The
dependency of the compressor power to the refrigerant mass flow rate is evident in Figure
93. Accordingly, the compressor power use decreased as the refrigerant mass flow rate
decreased (Figure 93).
8
8
7
7
6
6
5
5
4
4
3
3
2
2
1
Refrigerant Mass
Flow Rate (lb/min)
1
Avg Power = 2.24 kW
Avg Power = 1.93 kW
Avg Power = 2.12 kW
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
0
0
Test Scenarios
FIGURE 93.
COMPARISON OF TWO-MINUTE PROFILES OF COMPRESSOR POWER AND REFRIGERANT MASS FLOW RATE
OVER 24 HOURS – ALL THREE TEST SCENARIOS
Figure 94 depicts the 2-minute profile of condensate mass collected over the entire test
period for all three tested display cases. In Figure 94, the vertical lines indicate the mass of
melted frost during each defrost period and horizontal lines indicate the amount of moisture
collected during refrigeration periods. For the Hill Phoenix case, relatively flat horizontal
lines indicated an insignificant amount of collected moisture during refrigeration periods. For
the Hussmann and Tyler cases, on the other hand, the sloped horizontal lines indicate a
slightly higher amount of collected moisture during refrigeration periods.
More importantly, Figure 94 shows that the highest total mass of collected condensate over
a 24-hour period was for the Hill Phoenix case with 95.5 lbs. The lowest total mass of
collected condensate over a 24-hour period, however, was the Hussmann case with 76 lbs.
In other words, the amount of condensate collected over a 24-hour test period for the
Hussmann case was 20% less than the Hill Phoenix case and about 16% (76 lbs vs. 90.4
lbs) less than the Tyler case.
Page 78
June 2009
ET 06.07
100
90
Mass of Collected
Condensate (lbs)
80
Total Condensate = 95.5 lbs
Refrigeration Period
70
60
Defrost Period
50
Defrost Period
Defrost Period
40
30
20
10
0
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Test Scenarios
FIGURE 94.
COMPARISON OF TWO-MINUTE PROFILES OF MASS OF COLLECTED CONDENSATE OVER 24 HOURS – ALL
THREE TEST SCENARIOS
In general, the reduction in mass of collected condensate and RAT can be directly translated
to the reduction in infiltration load of the display cases. Figure 95 depicts that the infiltration
load was the lowest for the Hussmann display case (10,339 Btu/hr), and highest for the Hill
Phoenix case (13,881 Btu/hr) followed by the Tyler case (11,716 Btu/hr). This indicated
that the infiltration load of the Hussmann case was 26% lower than the Hill Phoenix case
and 12% lower than the Tyler case. Subsequently, since the infiltration load contributed to
about 80% of the total cooling load of these display cases, the total cooling load varied
according to variations in infiltration load.
Figure 95 also depicts the conduction, radiation, and internal load for all three display cases.
As shown, the conduction load was slightly higher for the Hill Phoenix case (637 Btu/hr)
when compared to the other two display cases (551 Btu/hr and 496 Btu/hr). This can be
attributed mainly to a slightly larger surface area of the case walls that are conducting heat.
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 lighting
system and evaporator fan motors, was slightly higher for the Hussmann case (730 Btu/hr)
when compared to the other two display cases (592 Btu/hr and 476 Btu/hr). This was
attributed mainly to an increase in evaporator fan motors power consumption for the
Hussmann case prior to initiation of defrost periods.
Page 79
June 2009
18,000
Conduction
Radiation
Internal (lights and evap fans)
Infiltration
12,621
Total
10,339
16,119
16,000
13,881
14,000
12,000
ET 06.07
13,662
11,716
10,000
8,000
6,000
4,000
2,000
1,010
637
1,001
592
973
730
551
496
476
0
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Test Scenarios
(tested display cases)
FIGURE 95.
COMPARISON OF TOTAL COOLING LOAD AND ITS COMPONENTS OVER 24 HOURS – ALL THREE TEST
SCENARIOS
Figure 96 compares the manufacturers published cooling load data with obtained test
results. The cooling loads shown are based on the linear-feet of the display case length. As
shown, in all instances the obtained cooling loads were higher than that specified by the
manufacturers. The highest percentage difference in the cooling load between the test and
published data was for the Tyler display case with 61%, and the lowest was for the
Hussmann display case with 15%.
2,500
Manufacturer Data
Total Cooling Load
per Linear-feet (Btu/hr/ft)
%  = 28%
2,000
Test Data
%  = 61%
2,015
%  = 15%
1,708
1,500
1,578
1,570
1,370
1,000
1,059
500
0
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Test Scnearios
FIGURE 96.
COMPARISON OF TEST DATA AND MANUFACTURER’S REPORTED COOLING LOAD PER LINEAR-FEET OF THE
DISPLAY CASE – ALL THREE TEST SCENARIOS
Figure 97 depicts the power usage by end-use and the total power for all three display cases
over the entire 24-hour test period. The case lighting power usage was similar for all three
display cases (0.11 kW to 0.12 kW). The evaporator fan power usage was highest for the
Page 80
June 2009
ET 06.07
Hussmann case (0.09 kW), and lowest for the Tyler case (0.03 kW). The Tyler display case
was the only case that was equipped with secondary or ambient fans that consumed about
0.02 kW. The lowest compressor power demand was observed for the Hussmann case with
1.93 kW, and the highest for the Hill Phoenix case with 2.24 kW followed by the Tyler case
with 2.12 kW. In other words, the compressor power demand for the Hussmann case was
14% less than the Hill Phoenix case and 9% less than the Tyler case. Subsequently, since
the compressor was the major contributor to the total display case power usage, the total
power varied according to variations in compressor power.
3.0
Evap. Fans
2.5
2.44
2.24
Power (kW)
2.0
Secondary Fans
Lighting System
Compressor
1.5
2.15
2.12
2.28
1.93
Total
1.0
0.5
0.06
0.0
0.11
n/a
Hill Phoenix
O5DM
0.09
0.12
n/a
Hussmann
M5X-GEP
0.03 0.02
0.11
Tyler
N6DHPACLA
Test Scenarios
FIGURE 97.
COMPARISON OF TOTAL AND END-USE POWER OVER 24 HOURS – ALL THREE TEST SCENARIOS
Figure 98 illustrates the total daily defrost duration and the refrigeration compressor run
time in terms of hours. Again, the defrost frequency and duration, as well as the defrost
method was in accordance with the case manufacturers specifications. The defrost duration
for the Hill Phoenix case was 2.8 hours, for the Hussmann case was 2.4 hours, and for the
Tyler case was 1.9 hours over a 24-hour period. Since the defrost method was off-cycle
where the compressor stops running, the refrigeration compressor run time or operation
hours was a function of defrost duration. Subsequently, when the defrost duration was low,
the compressor run time was high, and vice versa. The compressor operation hours for the
Hill Phoenix case was 21.2 hours, for the Hussmann case was 21.6 hours, and for the Tyler
case was 22.1 hours over a 24-hour period.
Page 81
June 2009
ET 06.07
28
Total Daily Defrost Duration and
Compressor Run Time (hours)
Daily Defrost Duration
24
Daily Compressor Run Time
20
22.1
21.6
21.2
16
12
8
4
0
2.8
Hill Phoenix
O5DM
2.4
Hussmann
M5X-GEP
1.9
Tyler
N6DHPACLA
Test Scnearios
FIGURE 98.
COMPARISON OF TOTAL DAILY DEFROST PERIODS AND REFRIGERATION (COMPRESSOR) RUN TIME OVER 24
HOURS – ALL THREE TEST SCENARIOS
The total daily energy and energy usage by end-use for all three display cases was captured
as well (Figure 99). It is important to note that over a 24-hour period of test runs the case
lighting system, evaporator fans, and the secondary fans were continuously on. The
refrigeration compressor was the only end-use component that its operation or run time
varied for each of the three tested display cases. The case lighting daily energy usage for
both the Hill Phoenix and Tyler cases was 3.0 kWh, and for the Hussmann case was 2.7
kWh. The evaporator fans daily energy usage was lowest for the Tyler case, 0.7 kWh, and
highest for the Hussmann case, 2.2 kWh. The daily energy usage of the secondary or
ambient fans of the Tyler case was about 0.4 kWh. The lowest compressor daily energy
usage was observed for the Hussmann case, 41.8 kWh, and the highest for the Hill Phoenix
case, 47.5 kWh, followed by the Tyler case, 46.9 kWh. Again, since the compressor was the
major contributor to the total display case energy usage, the total daily energy consumption
followed a similar pattern as the compressor daily energy usage.
Page 82
June 2009
ET 06.07
60
Evap. Fans
Daily Energy (kWh/day)
50
51.7
47.5
40
Secondary Fans
50.7
47.0
Lighting System
Compressor
46.9
41.8
Total
30
20
10
1.4
0
2.7
3.0
2.2
n/a
0.7
n/a
Hill Phoenix
O5DM
Hussmann
M5X-GEP
0.4
2.7
Tyler
N6DHPACLA
Test Scenarios
FIGURE 99.
COMPARISON OF TOTAL AND END-USE DAILY ENERGY OVER 24 HOURS – ALL THREE TEST SCENARIOS
Figure 100 illustrates the total cooling load and total power usage per refrigerated volume of
each display case. The refrigerated volume for each of the three display cases is also shown
in Figure 100. As shown, 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. A similar observation was made regarding the total power demand
per refrigerated volume. That is, 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.
200
Total Cooling Load per Refrigerated Volume (Btu/hr/ft³)
Total Cooling Load and Power
per Refrigerated Volume
180
160
175
Total Power per Refrigerated Volume (Watts/ft³)
150
140
147
120
100
80
60
Refrigerated
Vol. = 92.11 ft 3
Refrigerated
Vol. = 83.92 ft 3
Refrigerated
Vol. = 93.00 ft3
40
20
26.5
25.6
24.5
0
Hill Phoenix
O5DM
Hussmann
M5X-GEP
Tyler
N6DHPACLA
Test Scenarios
FIGURE 100. COMPARISON OF TOTAL COOLING LOAD AND POWER PER REFRIGERATED VOLUME – ALL THREE TEST
SCENARIOS
Page 83
June 2009
ET 06.07
Table 7 through Table 9 summarize and compare the key performance attributes of the Hill
Phoenix, Hussmann, and the Tyler display case. Table 7 summarizes and compares the key
system parameters and measured condensate mass for all three tested display cases. Table
8 summarizes and compares the key refrigeration parameters and cooling load for all three
display cases. Table 9 summarizes and compares power demand and energy consumption
for all three display cases.
TABLE 7.
SUMMARY OF KEY SYSTEM PARAMETERS AND MEASURED CONDENSATE FOR ALL THREE TEST SCENARIOS
KEY PERFORMANCE
ATTRIBUTES
HILL PHOENIX
(O5DM)
HUSSMANN
(M5X-GEP)
TYLER
(N6DHPACLA)
Discharge air
temperature (oF)
30.1
29.9
35.5
Return air temperature
(oF)
44.6
38.5
51.5
Suction pressure (psig)
57.7
61.4
58.9
Saturated evaporating
temperature (oF)
23.0
26.5
23.3
5.2
5.7
14.4
218.6
217.9
218.2
Saturated condensing
temperature (oF)
94.5
94.9
95.0
Total sub-cooling (oF)
29.9
31.4
62.1
Warmest product
temperature (oF)
41.8
39.8
48.4
Total measured
condensate over 24
hours (lbs)
95.5
76.0
90.4
Evaporator coil
superheat (oF)
Discharge pressure
(psig)
TABLE 8.
SUMMARY OF KEY REFRIGERATION PARAMETERS AND COOLING LOAD FOR ALL THREE TEST SCENARIOS
KEY PERFORMANCE
ATTRIBUTES
HILL PHOENIX
(O5DM)
HUSSMANN
(M5X-GEP)
TYLER
(N6DHPACLA)
Refrigeration effect
(Btu/lb)
57.6
58.5
58.4
Refrigerant mass flow
rate (lb/hr)
280
216
234
16,119
12,621
13,662
Total cooling load per
linear-feet (Btu/hr/ft)
2,015
1,578
1,708
Total cooling load per
refrigerated volume
(Btu/hr/ft3)
175
150
147
Conduction load (Btu/hr)
637
551
496
1,010
1,001
973
592
730
476
Total cooling load
(Btu/hr)
Radiation load (Btu/hr)
Internal load – case
lighting system and
evaporator fans (Btu/hr)
Page 84
June 2009
ET 06.07
KEY PERFORMANCE
ATTRIBUTES
HILL PHOENIX
(O5DM)
HUSSMANN
(M5X-GEP)
TYLER
(N6DHPACLA)
Infiltration load (Btu/hr)
13,881
10,339
11,716
21.2
21.6
22.1
2.8
2.4
1.9
Refrigeration run time
over 24 hours –
excluding defrost
(hrs/day)
Defrost duration over 24
hours (hrs/day)
TABLE 9.
SUMMARY OF POWER DEMAND AND DAILY ENERGY USAGE FOR ALL THREE TEST SCENARIOS
KEY PERFORMANCE
ATTRIBUTES
HILL PHOENIX
(O5DM)
HUSSMANN
(M5X-GEP)
TYLER
(N6DHPACLA)
Compressor power (kW)
2.24
1.93
2.12
Compressor daily energy
(kWh/day)
47.5
41.8
46.9
Evaporator fan motors
power (kW)
0.06
0.09
0.03
Evaporator fan motors
daily energy (kWh/day)
1.4
2.2
0.7
Secondary or ambient
fan motors power (kW)
n/a
n/a
0.02
Secondary or ambient
fan motors daily energy
(kWh/day)
n/a
n/a
0.4
Case lighting system
power (kW)
0.11
0.12
0.11
2.7
3.0
2.7
Total system power (kW)
2.44
2.15
2.28
Total system daily
energy (kWh/day)
51.7
47.0
50.7
Case lighting system
daily energy (kWh/day)
Page 85
June 2009
ET 06.07
CONCLUSIONS AND RECOMMENDATIONS
The results of this study indicated that the Hussmann display case had the lowest cooling
load requirement, and specifically infiltration load, when compared to the Hill Phoenix and
the Tyler display cases. This fact was traceable by low mass of condensate collected over a
24-hour period, and low temperature difference between the discharge and return air. Due
to low cooling load requirements, the refrigeration compressor power demand and daily
energy usage was the lowest for the Hussmann display case test scenario. This accordingly
resulted in lowering overall display case power demand and daily energy usage. Normalizing
the cooling load and power demand based on the refrigerated volume of the display cases
revealed that the performance of the Hussmann display case was better than the Hill
Phoenix case and to some extent comparable to, but not better than, the Tyler case.
One of the main reasons that the Hussmann display case had the lowest infiltration load was
attributed to its glass-front extension. The results revealed that by removing the glass-front
extension, the infiltration load increased by 28% and the compressor power increased by
18%. This finding was in line with a recent study that has identified the vertical distance
between the discharge and the return air grille to be one of the key geometric variables
impacting the infiltration load of open vertical refrigerated display cases [Ref. 4].
The results of this study also indicated that both the Tyler and the Hill Phoenix display cases
could not maintain the warmest product temperatures equal to or below 41oF, as required
by the Food and Drug Administrations’ (FDA) food code. The warmest product temperature
for the Tyler display case test scenario was around 48oF and for the Hill Phoenix case was
slightly below 42oF. The warmest product temperature for the Hussmann display case test
scenario, however, was slightly below 40oF.
Therefore, when selecting an open vertical refrigerated display case, it is recommended to
select a display case 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)
Page 86
June 2009
ET 06.07
REFERENCES
[1]
American Society of Heating, Refrigerating and Air-Conditioning Engineers,
Refrigeration Handbook, Chapter 46 – “Retail Food Store Refrigeration and Equipment,”
2006.
[2]
Baxter, V. D., “Investigation of Energy Efficient Supermarket Display Cases,”
ORNL/TM-2004/292. Oak Ridge national Lab, December 2004.
[3]
Itron. “California Commercial End-Use Survey: Consultant Report,” CEC-400-2006005. March 2006.
[4]
Southern California Edison. “Air Curtain Stability and Effectiveness in Open Vertical
Refrigerated Display Cases,” CEC 500-05-012. September 2008.
Page 87
June 2009

hussmann display case

Transcription

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