INVESTIGATING THE EFFECTS OF EVAPORATOR FAN SPEED VARIATION AND THE INFILTRATION PHENOMENON IN A WALK-IN COOLER

Size: px
Start display at page:

Download "INVESTIGATING THE EFFECTS OF EVAPORATOR FAN SPEED VARIATION AND THE INFILTRATION PHENOMENON IN A WALK-IN COOLER"

Transcription

1 Design & Engineering Services INVESTIGATING THE EFFECTS OF EVAPORATOR FAN SPEED VARIATION AND THE INFILTRATION PHENOMENON IN A WALK-IN COOLER Report Prepared by: Design & Engineering Services Customer Service Business Unit Southern California Edison August 03, 2011

2 Acknowledgements Southern California Edison s Design & Engineering Services (DES) group is responsible for this project. It was developed as part of Southern California Edison s Emerging Technology program under internal project number ET 10.SCE1050. DES project managers Sean Gouw and Rafik Sarhadian conducted this technology evaluation with overall guidance and management from Paul Delaney, and Ramin Faramarzi. Equally important was the great contribution and support from TTC staff Bruce Coburn, John Lutton, Grant Davis, Martin Vu and Scott Mitchell throughout this project. For more information on this project, contact either Sean.Gouw@sce.com or 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. Southern California Edison

3 ABBREVIATIONS AND ACRONYMS AHRI AHU ANSI ASHRAE Btu C C&S CEC CFD cfm CO 2 da DB DES DOE DP DR ECM EE F FDA fpm hp Air-Conditioning, Heating, and Refrigeration Institute Air Handling Unit American National Standards Institute American Society of Heating, Refrigerating ad Air-Conditioning Engineers, Inc. British thermal unit Celsius Codes and Standards California Energy Commission Computational fluid dynamics Cubic feet per minute Carbon dioxide Dry Air Dry-bulb (temperature) Design and Engineering Services Department of Energy (United States) Dew Point Demand Response Electronically commutated motor Energy Efficiency Fahrenheit Food and Drug Administration Feet-per-minute Horsepower Southern California Edison Page i

4 Hr HTTC HVAC IDSM K kpa kw kwh LDA LTTC M mm μm mj NDIR NIST PIV ppm RAT RH RTTC R-Value SCE SCT sec SET sq-ft hour HVAC Technology Test Center Heating, Ventilation, and Air Conditioning Integrated Demand Side Management Kelvin Kilo-Pascal Kilowatt Kilowatt hour Laser Doppler anemometry Lighting Technology Test Center Meter Millimeter Micrometer Millijoule Non-Dispersive Infrared National Institute of Standards and Technology Particle image velocimetry Parts per million Return Air Temperature Relative humidity Refrigeration Technology Test Center Thermal Resistance Value Southern California Edison Saturated condensing temperature Second Saturated evaporating temperature Square feet Southern California Edison Page ii

5 TAG TD TTC v V VSD Technical advisory group Temperature difference Technology Test Centers Water Vapor Volt Variable speed drive Southern California Edison Page iii

6 FIGURES Figure 1. Layout of the Walk-In Cooler Used Figure 2. Photograph of Evaporator Fan Coil of Walk-In Cooler Figure 3. Photograph of Refrigeration Compressors Figure 4. Figure 5. Figure 6. Figure 7. Figure 8. Figure 9. A) CO 2 Tracer Gas Tanks (left photo) B) Gas analyzer, Horiba VA-3000, (right photo) Location of Probes Corresponding to Second and Third Gas Analyzer Channels A) The Camera Mounts for the Sliding Door (left photo) B) The Laser and Mirror Mount (right photo) Quantity and Location of Sensors for Measuring Evaporator Discharge or Supply Air Dry-Bulb and Dew- Point Temperatures, Evaporator Fan Motor Power and Rotational Speed Quantity and Location of Sensors for Measuring Discharge or Supply Air Velocity Quantity and Location of Sensors for Measuring Evaporator Return Air Dry-Bulb and Dew-Point Temperatures, Evaporator Return Air Velocity, Refrigerant Temperature and Pressure Entering and Leaving the Evaporator Coil Figure 10. Quantity and Location of Sensors for Measuring Walkin s Interior and Exterior Dry-Bulb Temperatures and Relative Humidities, Interior and Exterior Walk-in Panel Surface Temperatures, and Product Temperatures Figure 11. Photograph of Filler Products and Temperature Poles Inside the Walk-In Figure 12. A) Photograph of Wall Surface Temperature Sensors (left photo) B) Photograph of a Product Simulator (right photo) Figure 13. Infiltration Rate as a Function of the Time Door Stands Open and Fan Speed For 75 F Dry-Bulb and 55% Relative Humidity Tracer Gas Method Figure 14. Infiltration Rate as a Function of the Time Door Stands Open and Fan Speed For 80 F Dry-Bulb and 60% Relative Humidity Tracer Gas Method Figure 15. Infiltration Rate as a Function of the Time Door Stands Open and Fan Speed For 84 F Dry-Bulb and 82% Relative Humidity Tracer Gas Method Southern California Edison Page iv

7 Figure 16. Infiltration Rate as a Function of the Time Door Stands Open and Fan Speed For 115 F Dry-Bulb and 14% Relative Humidity Tracer Gas Method Figure 17. The Natural Decay in the Cooler (closed door and stabilized system) as a Function of Fan Speed and Adjacent Space Dry-Bulb Temperature and Relative Humidity Figure 18. Infiltration Comparison for Sliding and Swing Type Doors at the Same Conditions (full speed fan and adjacent space condition of 75 F dry-bulb and 55% relative humidity) Figure 19. Infiltration Comparison of Strip Curtain and Propped Open Door/Inadequate Door Gasket Seal (adjacent space condition of 75 F dry-bulb and 55% relative humidity, and full fan speed) Figure 20. Infiltration Comparison between No Barrier and Strip Curtain with In/Out Traffic (adjacent space condition of 75 F dry-bulb and 55% relative humidity, and full fan speed) Figure 21. Infiltration Comparison between Different Door Positions (adjacent space condition of 75 F dry-bulb and 55% relative humidity, and full fan speed) Figure 22. Infiltration Rate as a Function of the Volume of Cold Air Inside the Walk-In Cooler Parametric Studies (adjacent space condition of 75 F dry-bulb and 55% relative humidity) Figure 23. Velocity Vectors or Profiles During the Infiltration Figure 24. Horizontal Velocity Contours During the Infiltration (Right +, Left -) Figure 25. Vertical Velocity Contours During the Infiltration (Up +, Down -) Figure 26. Dry-Bulb Temperature and Relative Humidity in Adjacent Spaces Figure 27. Key Controlled Variables A Look at Saturated Evaporating Temperatures and Saturated Condensing Temperatures Figure 28. Observing the Impacts of Varying Fan Speed on Product Temperatures for a Fixed Capacity Refrigeration System (product rack facing the fan coil east rack) Figure 29. Observing the Impacts of Varying Fan Speed on Product Temperatures for a Fixed Capacity Refrigeration System (product rack below the fan coil west rack) Figure 30. Observing the Impacts of Varying Fan Speed on Product Temperatures for a Fixed Capacity Refrigeration System (rack along the sliding door north rack) Southern California Edison Page v

8 Figure 31. Observing the Impacts of Varying Fan Speed on Product Temperatures for a Fixed Capacity Refrigeration System (rack along the swing type door south rack) Figure 32. Observing the Impacts of Varying Fan Speed on Walk- In Cooler s Interior Temperatures for a Fixed Capacity Refrigeration System (pole 1 near the swing type door) Figure 33. Observing the Impacts of Varying Fan Speed on Walk- In Cooler s Interior Temperatures for a Fixed Capacity Refrigeration System (pole 2 near the sliding door) Figure 34. Observing the Refrigeration Effect and Refrigerant Mass Flow Rate for a Fixed Capacity Refrigeration System Figure 35. Analyzing the Effects of Varying Fan Speed on Power Consumption for a Fixed Capacity Refrigeration System.. 58 Figure 36. Analyzing the Effects of Varying Fan Speed on Energy Consumption for a Fixed Capacity Refrigeration System.. 59 Figure 37. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack facing the fan coil east rack, scenario a) Figure 38. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack below the fan coil west rack, scenario a) Figure 39. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack along the sliding door north rack, scenario a) Figure 40. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack along the swing type door south rack, scenario a) Figure 41. Observing the Walk-In Cooler s Interior Temperatures for a Variable Capacity Refrigeration System (pole 1 near the swing type door, scenario a) Figure 42. Observing the Walk-In Cooler s Interior Temperatures for a Variable Capacity Refrigeration System (pole 2 near the sliding door, scenario a) Figure 43. Observing the Refrigeration Effect and Refrigerant Mass Flow Rate for a Variable Capacity Refrigeration System (scenario a) Figure 44. Observing the Power Consumption Trends for a Variable Capacity Refrigeration System (scenario a) Figure 45. Observing the Energy Consumption for a Variable Capacity Refrigeration System (scenario a) Figure 46. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack facing the fan coil east rack, scenario b) Southern California Edison Page vi

9 Figure 47. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack below the fan coil west rack, scenario b) Figure 48. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack along the sliding door north rack, scenario b) Figure 49. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack along the swing type door south rack, scenario b) Figure 50. Observing the Walk-In Cooler s Interior Temperatures for a Variable Capacity Refrigeration System (pole 1 near the swing type door, scenario b) Figure 51. Observing the Walk-In Cooler s Interior Temperatures for a Variable Capacity Refrigeration System (pole 2 near the sliding door, scenario b) Figure 52. Observing the Refrigeration Effect and Refrigerant Mass Flow Rate for a Variable Capacity Refrigeration System (scenario b) Figure 53. Observing the Power Consumption Trends for a Variable Capacity Refrigeration System (scenario b) Figure 54. Observing the Energy Consumption for a Variable Capacity Refrigeration System (scenario b) Figure 55. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack facing the fan coil east rack, scenario c) Figure 56. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack below the fan coil west rack, scenario c) Figure 57. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack along the sliding door north rack, scenario c) Figure 58. Observing the Product Temperature Trends for a Variable Capacity Refrigeration System (rack along the swing type door south rack, scenario c) Figure 59. Observing the Walk-In Cooler s Interior Temperatures for a Variable Capacity Refrigeration System (pole 1 near the swing type door, scenario c) Figure 60. Observing the Walk-In Cooler s Interior Temperatures for a Variable Capacity Refrigeration System (pole 2 near the sliding door, scenario c) Figure 61. Observing the Refrigeration Effect and Refrigerant Mass Flow Rate for a Variable Capacity Refrigeration System (scenario c) Figure 62. Observing the Power Consumption Trends for a Variable Capacity Refrigeration System (scenario c) Southern California Edison Page vii

10 Figure 63. Observing the Energy Consumption for a Variable Capacity Refrigeration System (scenario c) Figure 64. Impacts of Floating Head Pressure on Saturated Evaporating Temperature and Saturated Condensing Temperature for a Fixed Capacity Refrigeration System.. 90 Figure 65. Impacts of Floating Head Pressure on Power Consumption for a Fixed Capacity Refrigeration System.. 91 Figure 66. Impacts of Floating Head Pressure and Variable Speed Evaporator Fan Controls on Power Consumption for a Fixed Capacity Refrigeration System Figure 67. Impacts of Floating Head Pressure on Energy Consumption for a Fixed Capacity Refrigeration System.. 93 Figure 68. Impacts of Floating Head Pressure and Variable Speed Evaporator Fan Controls on Energy Consumption for a Fixed Capacity Refrigeration System Figure 69. Impacts of Floating Head Pressure on Saturated Evaporator Temperature and Saturated Condensing Temperature for a Variable Capacity Refrigeration System Figure 70. Impacts of Floating Head Pressure on Power Consumption for a Variable Capacity Refrigeration System Figure 71. Impacts of Floating Head Pressure and a Variable Speed Evaporator Fan on Power Consumption for a Variable Capacity Refrigeration System Figure 72. Impacts of Floating Head Pressure on Energy Consumption for a Variable Capacity Refrigeration System Figure 73. Impacts of Floating Head Pressure and a Variable Speed Evaporator Fan on Energy Consumption for a Variable Capacity Refrigeration System Southern California Edison Page viii

11 TABLES Table 1. Summary of Test Scenarios Table 2. Specifications and Calibration Dates for Sensors Used Table 3. Table 4. Table 5. Table 6. Summary of Maintained Adjacent Space Conditions for all Fixed Capacity Test Runs Summary of Average, Minimum, and Maximum Product Temperatures for all Fixed Capacity Test Runs Average Saturated Evaporating/Condensing Temperatures, Average Product Temperatures, and Average Adjacent Space Conditions for Variable Capacity Test Runs Average Saturated Evaporating/Condensing Temperatures, Average Product Temperatures, and Average Adjacent Space Conditions for Variable Capacity Test Runs (Continued) Southern California Edison Page ix

12 EQUATIONS Equation 1. Tamm s Equation for Air Infiltration Rate... 6 Equation 2. Gosney and Olama s Air Exchange Equation for Fully Established Flow... 7 Equation 3. Cleland s Equation for Air Infiltration Rate... 7 Equation 4. Finding Infiltration Using Original Cold Air Exfiltration Determined from the Tracer Gas Technique... 9 Equation 5. Fan Laws Equation 6. Airflow Rates Inside Walk-in Equation 7. Refrigeration Effect Equation 8. Total Refrigeration Load (in btu/hr) Equation 9. Total Refrigeration Load (in cooling tons) Equation 10. Volumetric Air Flow Rate Equation 11. Evaporator Coil Superheat Equation 12. Condenser Coil Sub-cooling Equation 13. Total Refrigeration Power Usage, Excluding Condenser Equation 14. Energy Usage Equation 15. Conduction or Transmission Load of the Walk-In Southern California Edison Page x

13 CONTENTS EXECUTIVE SUMMARY 1 INTRODUCTION 3 Definition, Purpose and Application of Walk-Ins... 3 Energy Usage of Walk-Ins Refrigeration System... 3 Project Rationale... 4 BACKGROUND 6 Common Methods for Estimating Infiltration... 6 Tracer Gas Technique... 8 Relationship between Performance Variables for Fans ASSESSMENT OBJECTIVES 11 TECHNICAL APPROACH 12 WALK-IN COOLER DESCRIPTION 18 TEST METHODOLOGY 20 Infiltration Rate Study Test Design and Instrumentation Varying Evaporator Fan Speed Study Test Design and Instrumentation DATA ANALYSIS 30 RESULTS 34 Results of Infiltration Rate Study Results of Varying Evaporator Fan Speed Study CONCLUSIONS 98 RECOMMENDATIONS 99 APPENDIX A CALCULATIONS OF MIXTURE DENSITY OF AIR 100 APPENDIX B NATURAL INFILTRATION THROUGH UNSEALED AREAS 102 APPENDIX C FINDING TRACER GAS CONCENTRATION GRADIENT IN TIME 103 Southern California Edison Page xi

14 APPENDIX D CFD ANALYSIS 106 APPENDIX E INFILTRATION RATES FOR VARIOUS ADJACENT SPACE CONDITIONS 108 APPENDIX F TECHNOLOGY TEST CENTERS 120 Refrigeration Technology Test Center Responsibilities Test Chambers and Equipment REFERENCES 123 Southern California Edison Page xii

15 EXECUTIVE SUMMARY This Emerging Technology assessment project was conducted at Southern California Edison s Technology Test Centers controlled environment chambers and focuses on two main goals: The first is to investigate the impact of varying evaporator fan speed on product temperature and energy efficiency of a walk-in cooler, the second is to investigate the infiltration phenomenon in a walk-in cooler. Two key factors led to this project. First, there is no feasibility study addressing the impact of varying airflow rates on product quality and safety. Second, there are limited and non-substantial empirical evidence with regards to the infiltration phenomenon in walk-ins. Both are crucial factors to consider since infiltration of warm air into walk-ins contributes to increases in cooling load and energy consumption of refrigeration systems. This project addresses challenges associated with implementation of energy efficiency strategies in walk-ins and compliance with the food safety regulations and quality. Evaluating the impact of varying evaporator fan speed with respect to demand reduction, energy savings, and product temperature will establish the feasibility of reducing fan speed, and hence airflow rates, in the walk-ins. This is not only with regards to the improvements in refrigeration system efficiency, but also with respect to the ability of maintaining desirable product temperatures. This evaluation encompasses the quantification of direct and interactive effects of reduced fan speed on power and energy usage of evaporator fan motor and refrigeration compressors. More importantly, the effects of reduced fan speed, hence airflow rates inside the walk-in, on maintaining proper and desired product temperatures are investigated. These evaluations were conducted for the two most common refrigeration systems (single fixed capacity compressor and multiplex variable capacity compressor rack) under three ambient conditions. Quantification and identification of factors contributing to the infiltration in walk-ins will ultimately pinpoint to design and operational advances for minimizing infiltration and increasing overall refrigeration system efficiency. This project quantifies the infiltration rate of warm and moist air from the adjacent space into the walk-in cooler. The results obtained were used to evaluate the validity of currently practiced methodologies for estimating the infiltration rate. Factors that affect the infiltration rate are also examined. These factors included door opening duration, evaporator fan speed, air dry-bulb temperature and relative humidity of adjacent space, presence of strip curtain at the doorway, propped open door/inadequate door gasket seal, and door type (hinged versus sliding). These tests will be accomplished by using the tracer gas technique with the main door closed and opened for various periods of time. The tracer gas technique directly measures the loss of original cold air in the walk in cooler for each prescribed scenario. These measurements are used with fundamental equations (e.g., conservation of mass) to determine infiltration. Computational fluid dynamics analysis is also performed to identify temperature variations inside the cooler for proper positioning of tracer gas sampling probes. In addition, the particle image velocimetry (PIV) technique is used to map the velocity profiles at the doorway due to the infiltration. The following lists the conclusions drawn from this project: The infiltration rate is a time-dependent process that follows a decaying function behavior as conditions between air inside the walk-in and air in adjacent space reach equilibrium. Before coming to equilibrium, 75% of infiltration occurs (75% of cold air Southern California Edison Page 1

16 in the cooler is lost) during the first 30 seconds of a door opening. Equilibrium is reached (all cold air in the cooler is lost) at 120 seconds of a door opening. The infiltration rate is also dependent on how much product the walk-in cooler contains. An empty cooler has higher volumes of cold air that may be lost. Modeling results for this cooler show that initial infiltration rates are approximately 4 ft 3 /sec for a 25% empty cooler, 17 ft 3 /sec for a 50% empty cooler, 38 ft 3 /sec for a 75% empty cooler, and 68 ft 3 /sec for a 100% empty cooler. The current industry-wide methodologies for estimating the infiltration rate in their steady state form are not sufficient for predictions. They over-estimate the infiltration rate for durations of more than seconds. Even after modifying for time-varying densities, they still over-predicted infiltration and the cooling load thereafter. The fan speed does not impact the infiltration rate for small temperature differences between the refrigerated space and adjacent warm space. During the period of door opening, for a temperature difference of about 45 F turning the fan off from running at full speed will lower the infiltration by about 5-7%. However, as this temperature difference increases to 80 F, the reduction in infiltration will be about 15-20% during the first 20 seconds of the infiltration process. The flow visualization technique of PIV mapped the velocity profile and proved that the schematics currently used to demonstrate air movement at the door opening capture the direction of motion. However, the values of the discharge velocity components have never been measured by the state-of-the-art PIV technology. There were no significant changes or impacts on the product temperature maintenance as well as room temperature stratification as a function of fan speed. The cooling load was the most dominant factor in power demand and energy usage compared to evaporator fan speed variations. This indicated that the heat gain from the ECM fan motor had little to no impact on the total cooling load due to its low power consumption. Floating heat pressure had impact that is more substantial on power demand and energy consumption than the variable evaporator fan speed. This was observed in both fixed and variable capacity refrigeration systems. To understand the full impact of individual parameters on infiltration, such parameters must be adjusted singularly. Different fan placement is recommended as a new parameter that needs to be tested with computational fluid dynamic technique and a modular scaled model. Although the model developed in this project links the volume of air inside the cooler to the infiltrated amount of air, more tests with several filled volumes of the cooler are recommended for further verification. To confirm that there are no significant differences between the door type (swing vs. sliding) on the infiltrate rate, testing at the sliding door needs to be duplicated for swing or hinged type doors. Further verification of this claim may be done through the use of the PIV technique at the midway plane of the swing type door. Additionally, to confirm the impacts of variable speed evaporator fan motors in walk-ins, it is suggested to further investigate the fan speed variations in a walk-in freezer applications as well as larger walk-in sizes and configurations. Southern California Edison Page 2

17 INTRODUCTION This laboratory assessment project studies the feasibility of reducing evaporator fan speed as a function of cooling load or heat gain variations inside a walk-in cooler. This encompasses the quantification of direct and interactive effects of reduced fan speed on power demand and energy usage of evaporator fan motor and refrigeration compressors. More importantly, the effects of reduced fan speed, hence airflow rates inside the walk-in, on maintaining proper and desired product temperatures are investigated. These evaluations were conducted for the two most common refrigeration systems (single fixed capacity compressor and multiplex variable capacity compressor rack) under three ambient temperatures. This project also investigates the infiltration rate of warm and moist air from the adjacent space into the walk-in cooler. The results obtained were used to evaluate the validity of currently practiced methodologies for estimating the infiltration rate. Factors that affect the infiltration rate were also examined. These factors included door opening time, evaporator fan speed, air temperature and relative humidity of adjacent space, strip curtain at the doorway, propped open doors/inadequate door gasket seal, and door type (hinged versus sliding). These were accomplished by means of tracer gas technique with the main door being closed and opened for various periods of time. Particle image velocimetry (PIV) technique was also used to map the velocity profiles at the doorway due to the infiltration. DEFINITION, PURPOSE AND APPLICATION OF WALK-INS Walk-ins are defined as room sized compartments that achieve controlled storage conditions using thermal insulation and refrigeration equipment. By definition, the floor area of a walk-in is equal to or less than 3,000 square feet (sf). Walk-ins are used to provide short term storage for perishable food products to maximize food safety and shelf life. Walk-ins are classified either as coolers or freezers. Walk-in coolers are used for medium-temperature (above 32 F) applications such as fresh fruits and melons, fresh vegetables, fresh meat and dairy products. Walk-in freezers are used for low-temperature (below 32 F) applications such as frozen packed fruits and vegetables, frozen meat and dairy products. Walk-ins are typically found in restaurants, both fast food and sit-down restaurants, as well as small and medium to large grocery stores or supermarkets. Small grocery store refers to convenience stores and independently owned small food markets. ENERGY USAGE OF WALK-INS REFRIGERATION SYSTEM According to California Energy Commission (2000), refrigeration contributes to about 20% of total energy usage in restaurants, and to about 38% of total energy usage in supermarkets and grocery stores. This estimation is in agreement with Southern California Edison s (SCE s; 1996, 1999) Technology Test Centers (TTC) energy audit data conducted for two major chain supermarket customers. According to this data, depending on the supermarkets size and layout, the refrigeration can contribute somewhere between 35% and 55% of total energy usage. In addition, up to 20% of the refrigeration energy usage is due to the walk-ins. On average, the annual energy Southern California Edison Page 3

18 usage of combination of walk-in coolers and freezers is about 330 kwh per sq-ft (Southern California Edison, 2007). PROJECT RATIONALE What prompted the initiation of this project is attributed to two facts about walk-ins. Here are the broad issues, but each one is discussed in greater detail in the following paragraphs. First, there is a lack of empirical studies that used robust techniques and methodologies to precisely and directly measure the infiltration rate into the walkins. Second, there is a lack of feasibility studies investigating the impact of varying the evaporator fan speed on product quality and safety. Infiltration rate refers to the entrainment of warm and moist air from the adjacent spaces into the cold or refrigerated spaces such as walk-ins. Infiltration can constitute more than half of the total heat gain or cooling load (Pham and Oliver, 1983). Infiltration is a major factor in the design, operation, and performance of refrigeration systems. In essence, infiltration of warm air into the walk-in coolers and freezers is mainly responsible for the total cooling load or heat gain and increased energy consumption of refrigeration compressors thereafter. Frequent opening and closing of doors, and the duration for which they are left open can contribute to the infiltration load of walk-ins. Infiltration can also occur through the small openings and unsealed or poorly sealed areas. Here, this process is referred to as the natural infiltration. To minimize the infiltration during door opening periods, typically infiltration barriers such as strip curtains or double swing type doors are used. Thus far, however, there has not been any study that employed robust methodology and technology to precisely and directly measure the natural infiltration or infiltration during the door opening periods. Systematic and direct measurement of the infiltration rate would allow the identification of operating conditions that could reduce the infiltration rate. It will also provide empirical data to assess the validity of widely used equations by industry and designers for estimating the infiltration load of walk-ins. In addition, there has not been any flow visualization of the infiltration or discharge process mapping the velocity profiles at the doorway. Mapping the velocity profiles, including the magnitude and direction of velocity, will enhance understanding of the air exchanges that takes place when the door is being opened to the adjacent space. Further, from operational standpoint, evaporator fan motors of walk-in coolers and freezers run continuously year-round. In other words, regardless of the heat gain or cooling load requirements of walk-ins, the evaporator fan motors operate constantly at their maximum speed. This is also true during defrost periods for the walk-in coolers that rely on off-cycle defrost for melting the ice build-up on the coils. So, the opportunity for energy efficiency improvements is to reduce the speed of the evaporator fan according to the cooling load variations. For instance, when the refrigeration system is able to achieve and maintain walk-in temperature set point, the fan speed can be reduced. Since the fan power is a cubic function of the fan speed, a small reduction in speed would result in reduced fan power demand and energy usage. Reduced fan power, conceptually, will result in reduced heat gain from the fan into the walk-in, thereby reducing overall cooling load and refrigeration compressors energy consumption. More importantly, since the fan speed is directly proportional to the airflow rate, reducing fan speed will reduce the air circulation or airflow inside the walk-in. The Southern California Edison Page 4

19 impact of reduced airflow rates inside the walk-in due to reduced fan speed has not been evaluated and documented. Thus, although it is well understood that reducing the fan speed will reduce power demand and energy usage, its impact on product temperatures stored in the walk-in due to reduced airflow rates is not a well understood phenomenon. Since product safety and quality is absolutely the most important factor in walk-ins design and operation, it becomes imperative to evaluate the impact of fan speed in terms of maintaining desired product temperatures. Subsequently, methodologies and protocols were developed and used to directly measure the infiltration rate using concentration of carbon dioxide (CO 2 ) tracer gas. Results obtained were used to examine the validity of three most common methods used by industry and designers for estimating the infiltration load of walk-ins. This project also examined the correctness of the speculations about the velocity profile during the infiltration via PIV flow visualization technique. Additionally, the impact of varying evaporator fan speed on product quality, power demand and energy usage was examined under various cooling load and outdoor ambient conditions. This was done for the two most common refrigeration systems, namely single and multiplex compressor systems. Southern California Edison Page 5

20 BACKGROUND This section starts with three common methodologies that are currently used by industry and designers for estimating infiltration rate and infiltration load in walk-in coolers and freezers. Following that, brief discussions on the tracer gas and particle image velocimetry (PIV) techniques are provided. This section ends with a discussion on the relationship between performance variables for fans. COMMON METHODS FOR ESTIMATING INFILTRATION The infiltration phenomenon (when the walk-in cooler door is open) can be described as the colder denser air spilling or exfiltrating out, while the ambient outside air infiltrates into the cooler. Based on mass conservation law, the amount of warm air infiltrated into the cooler and the amount of cold air exfiltrated out of the cooler are equal. Thus, in this report the terms infiltration and exfiltration are used interchangeably and synonymously. There is a common equation that is frequently used to estimate the infiltration rate of warm air into a cold room through doorways. This equation, (Equation 1), was derived by Tamm (1965). It correlates the infiltration rate to geometric dimensions of the door, and densities of warm and cold air. The effect of relative humidity is embedded in the density of air. EQUATION 1. TAMM S EQUATION FOR AIR INFILTRATION RATE ( ) 2WH 2gH 1 - s Q = s where, 1/3 ( ) Tamm 3 Q Tamm = air Infiltration rate, ft 3 /sec W H = door width, ft = door height, ft g = gravitational acceleration, ft/ sec 2 s = ratio of warm air density, ρ outside (lb m /ft 3 ), to cold air density, ρ inside (lb m /ft 3 ), dimensionless This equation has been modified by Chen et al. (2002) and East et al. (2003) for the impact of door open time, plastic strip curtains, and traffic through doorways. However, this equation is based on a fully developed flow between the warm and cold areas. The most recent work is presented by Reindl and Jekel (2009) where they used carbon dioxide as a tracer gas in a blast freezing environment. The infiltration takes place naturally through small openings and was measured by an infrared hand-held detector. Another widely used equation is Gosney and Olama s (1975) air exchange equation. It is a derivation of the cooling load as a function of the geometry of the door and Southern California Edison Page 6

21 density and enthalpy of the warm and cold air. In fact, this equation is suggested to be used for estimating the infiltration load of walk-ins in American Society of Heating, Refrigerating and Air-Conditioning Engineers (ASHRAE) 2006 Refrigeration Handbook. This equation does not directly calculate the infiltration rate, but it links the infiltration rate, through an energy balance equation, to the cooling load for fully developed flows. Equation 2 shows Gosney and Olama s air exchange equation for fully established flow (ASHRAE, 2006). EQUATION 2. GOSNEY AND OLAMA S AIR EXCHANGE EQUATION FOR FULLY ESTABLISHED FLOW é ù 0.5 æ ö ( ) ( ) ρoutside q = A h outside - hinside ρinside 1 - ( gh) çè ø 1/3 ρ æ ö inside ρoutside 1+ ç êë çè ρ inside ø úû where, q = sensible and latent refrigeration load, Btu/hr A = doorway area, ft 2 h outside = outside or adjacent space air enthalpy (infiltrated air), Btu/lb h inside = inside or refrigerated space air enthalpy, Btu/lb ρ outside = outside or adjacent space air density (infiltrated air), lb/ft 3 ρ inside = inside or refrigerated space air density, lb/ft 3 g = gravitational acceleration, ft/sec 2 H = door height, ft 1.5 Another approach to quantify air infiltration based on hydrodynamic theory for a flow through an orifice is Cleland s equation (Equation 3). However, Equation 3 does not take into account the abrupt expansion. To do this, a correction factor needs to be multiplied. EQUATION 3. CLELAND S EQUATION FOR AIR INFILTRATION RATE 2ΔP Q Cleland = CoA ρ inside where, Q Cleland = air Infiltration rate, ft 3 /sec C o = orifice coefficient, dimensionless A = door area, ft 2 D t = door open factor, dimensionless P = pressure difference between outside and inside of the cooler, lb/ft 2 ρ inside = inside or refrigerated space air density, lb/ft 3 Southern California Edison Page 7

22 The validity of these equations has not been examined if the infiltration process is transient. Subsequently, Tamm s, Gosney s and Cleland s equations were examined in this project. However, special attention should be given to the term of type: æ ö ρoutside - ç 1 çè ρ ø inside If the values of inside and outside temperatures are fixed at a prescribed value, both of these equations will yield a constant value of infiltration (and simultaneous exfiltration) that is somewhat far from the reality. For this assumption to be true, both rooms must be considered as reservoirs, which is not the case. However, if the cooler air density is taken to be a function of time as infiltration dictates, a more realistic estimate of infiltration by these two equations may be obtained. This requires information about the average air temperature in the cooler as a function of time after the air inside the cooler stabilizes. In the current experiment this stabilization of inside air could be easily detected by the tracer gas, enabling to find this functionality and use a transient value for air density inside the cooler to present modified Tamm, Gosney or Cleland equations. Using hot-wire anemometers and thermocouples as described by Hendrix et al. (2008) is another method to determine the infiltration rate. However, the setup often provides limited real time measurements. Foster, et al. (2002) have used computational fluid dynamics (CFD) and laser Doppler anemometry (LDA) to quantify the amount of infiltrated air by finding the velocity profiles. This method is cumbersome and not a very practical method to measure the infiltration rate due to its excessive time requirement. A more practical method is the use of a CO 2 tracer gas with a gas analyzer to monitor the concentration inside the cooler in real time. This method has been used by Amin, et al. (2009) for measuring the infiltration rate of open refrigerated display cases and is adopted in this work as a robust, accurate and simple-to-use technique. TRACER GAS TECHNIQUE There is a wide range of applications for tracer gas technique. The application of interest here is the use of tracer gas to measure the amount of infiltration in walkins. A tracer gas is a substance that is used to tag volumes of air in order to be able to infer their bulk movement. In essence, the exchange of air can be quantified by monitoring the injection and concentration of the tracer gas. In this project CO 2 was used as the tracer gas because it is cost effective and safer compared to other substances. Subsequently, to find the infiltration rate using tracer gas technique, the variation of CO 2 concentration in time was obtained using Equation 4. Southern California Edison Page 8

23 EQUATION 4. FINDING INFILTRATION USING ORIGINAL COLD AIR EXFILTRATION DETERMINED FROM THE TRACER GAS TECHNIQUE a Q= " dt Q - m OR = " OR = ρ OR dc mix dc " m= -ρmix CO2 dt " dc dt a dc CO2 dt Given that, m = ρ Q mix ρ = ρ C + ρ C + ρ C mix da da v v CO2 CO2 C a = 1 - C dc a =- dt dc CO2 CO2 dt where, Q = infiltration volumetric flow rate, as determined with tracer gas measurements (losses of original cold air) ft 3 /sec m = infiltration mass flow rate, as determined with tracer gas measurements (losses of original cold air) lb/sec ρ mix = total effective gas mixture density within the walk-in cooler, lb/ft 3 " = volume of the room, ft 3 C a C CO2 t = mass fraction of humid air, dimensionless = mass fraction of carbon dioxide, dimensionless = time, sec ρ da = density of dry air, lb/ft 3 C da = mass fraction of dry air, dimensionless ρ v = density of water vapor, lb/ft 3 C v = mass fraction of water vapor, dimensionless ρ CO2 = density of carbon dioxide, lb/ft 3 Southern California Edison Page 9

24 Average volumetric infiltration rates are found by using the door opening time and the decay in CO 2 concentration from the beginning to the end of the test period. Infiltration mass flow rates may be found by incorporating the gas mixture density within the walk-in cooler. Although the tracer gas concentration is relatively small and had minor effects on the air mixture density, it was included in time varying density calculations for precision. RELATIONSHIP BETWEEN PERFORMANCE VARIABLES FOR FANS Almost all walk-in coolers and freezers use forced-circulation evaporators with propeller type fans powered by fractional horsepower (hp) motors. Evaporator units are integral component of walk-in coolers and freezers refrigeration system. The fans housed within these units continuously move air across the evaporator coils and are responsible for circulating air throughout the refrigerated space. Three most common evaporator fan motor types are the shaded-pole, permanent split capacitor and electronically commutated motor (ECM). The shaded-pole and permanent split capacitor motors are induction type and are powered by alternating current electricity. A variable frequency drive controller is used to vary the speed of these types of motors by varying the frequency to the motor. ECMs, on the other hand, are brushless synchronous electric motors and are powered by direct current electricity. A variable speed drive controller is used to vary the speed of these types of motors by varying the voltage to the motor. Since ECMs are more efficient than the shaded-pole and permanent split capacitor type motors, the current regulations mandate using ECMs in the walk-ins. Therefore, in this project ECM motors were used. The relationship between performance variables (airflow rate, rotational speed, pressure and power) of the fans are governed by fan laws as shown in Equation 5 (ASHRAE, 2004). The subscripts 1 and 2 denote states or conditions. The fan law describes that the volumetric airflow rate is directly proportional to the rotational speed. The pressure and power, on the other hand, are square- and cube-function of the speed and airflow rate, respectively. EQUATION 5. FAN LAWS 1/2 1/3 Q æ ö æ ö 1 N1 P1 W1 = = = ç ç Q N èp ø èw ø where, Q = volumetric airflow rates, ft 3 /min (cfm) N = rotational speed, revolutions per minute (rpm) P = pressure, inches of water W = power, watts Southern California Edison Page 10

25 ASSESSMENT OBJECTIVES The main and overall objectives of this project are to enhance understanding about the infiltration phenomenon in walk-ins and address the unknowns associated with varying airflow rates in walk-ins in terms of product quality and safety. The specific objectives include; 1. Evaluate the operational parameters including thermal load of walk-ins 2. Quantify infiltration rate and the amount of infiltrated air using CO 2 tracer gas technique as a function of: a. door opening duration b. adjacent space temperature variations c. infiltration barriers such as strip curtains d. fan speed variations e. door open area f. propped open door/inadequate door gasket seals 3. Demonstrate the infiltration pattern at the doorway using PIV visualization technique 4. Evaluate the validity of current methodologies for estimating infiltration rate into walk-ins 5. Quantify power demand reduction and energy savings associated with varying the evaporator fan speed based on thermostatic setpoint for both single and multiplex compressor systems 6. Evaluate the impact of varying fan speed on product temperature for both single and multiplex compressor systems Southern California Edison Page 11

26 TECHNICAL APPROACH Below is the summary of overall project tasks. The details of each task are described in the subsequent paragraphs. Task 1: Task 2: Task 3: Task 4: Task 5: Task 6: Task 7: Task 8: Task 9: Identify project team and stakeholders Review the current walk-in test standards Review literature on the infiltration studies in walk-ins Laboratory testing of ECM fan with and without variable speed drive Fabricate CO 2 tracer gas probes Measure infiltration rate using CO 2 tracer gas technique Identify the impact of varying fan speed on infiltration rate Identify the impact of temperature variations between the walk-in and the adjacent space on infiltration rate Identify the impact of door opening duration, strip curtains, propped open door/inadequate door gasket seals, door types (swing vs. sliding), and door open area on infiltration rate Task 10: Map the velocity profile of infiltrated warm air and exfiltrated cold air Task 11: Data Analysis Task 12: Prepare the final report Task 1: Identify project team and stakeholders This task identifies and establish technical advisory group (TAG) and continuously be in communication with them on project-related issues. This ensures the robustness and usefulness of the finalized test plan and scenarios including monitoring plan. Task 2: Review the current walk-in test standards This task gathers and reviews all the relevant standards and documentation to establish test scenarios and monitoring plan for laboratory testing. The major sources for obtaining such information are: (1) Air-Conditioning, Heating, and Refrigeration Institute (AHRI), (2) American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (ASHRAE), and (3) state and federal rule making bodies. Task 3: Review of literature on the infiltration studies in walk-ins This task gathers and reviews all existing literature and research that focuses on the infiltration of walk-ins. Understanding the existing procedures and approaches, and possibly their weaknesses, ultimately guides and directs the efforts of this project in terms of enhancing analytical and experimental procedures and methods. Task 4: Laboratory testing of ECM fan with and without variable speed drive This task quantifies power demand reduction and energy savings associated with varying the speed of evaporator ECM fan. In addition, monitor product temperatures to examine the temperature variations, if any, as a result of varying fan speed. This is accomplished by comparing the results of testing ECM with and without VSD in a laboratory setting. Southern California Edison Page 12

27 The following sub-tasks summarize how the main goal of this task was achieved. The goals and objectives of sub-tasks 4.1, 4.2, and 4.3 were to ensure walk-in was setup properly to resemble realistic conditions, collect power and energy usage data of the refrigeration system as well as the evaporator fan motor without and with VSD, respectively. The data gathered was used to compare the demand and energy savings associated with installing VSD on the ECM. In addition, product temperatures and temperature gradient inside the walk-in were measured and recorded. This established the impact of using VSD on ECM in walk-ins not only from the power and energy standpoint, but also from product quality and safety stance. Task 4.1: Identify proper walk-in setup in terms of: airflow rates cooling load conditions walk-in cooler temperature and humidity adjacent space temperature and humidity dummy or filler products and test or product simulators defrost frequency (test duration) Task 4.1: Measure power demand and energy usage of the refrigeration system and fan motor when ECM is not equipped with VSD Task 4.2: Measure power demand and energy usage of the refrigeration system and fan motor when ECM is equipped with VSD Table 1 summarizes the test scenarios. All test scenarios were conducted without using strip curtains. These test scenarios were conducted to quantify the effects of floating head pressure and variable speed evaporator fans for fixed capacity and variable capacity rack systems. Three levels of saturated condensing temperatures were incorporated per test scenario to capture the effects of floating head pressure. Several scenarios were explored to capture the effects of variable speed evaporator fans. Savings associated with fixed capacity systems were examined by quantifying the energy saved as a result of stepping down fan speed during compressor OFF cycles. In addition, various levels of evaporator load were used to compare load variations. For rack system, the refrigeration capacity was modulated by varying the compressor speed. For every scenario associated with variable capacity rack system, two levels of evaporator load were established to demonstrate the benefits of VSD fans at part load versus full load. Although, initially, it was anticipated to investigate floating suction pressure through fan speed modulations, this was later not considered. The reason for not pursuing this was because during testing it was observed and determined that the heat gain from the ECM fan was insignificant and too minimal to affect the suction pressure. Southern California Edison Page 13

28 TABLE 1. SUMMARY OF TEST SCENARIOS Test # 1 System Type Evaporator Load Test Scenarios Compressor Control Strategy Suction Pressure Evaporator Fan SCT ( F) 2 High* 100% Fixed Compressor 70 5 High* speed 70% Fixed 90 Variable 6 modulates to Capacity 110 match Fixed 7 Compressor evaporator 70 (Rack System) 8 Low** load, no 100% Fixed 90 9 cycling Low** 70% Fixed Fixed Capacity 14 15% Oversized capacity 100% Fixed 90 Compressor 15 (Fixed Output 110 High* to peak load, Fixed 16 Remote Stepped control, 70% 70 fixed speed, 17 Condensing during compressor cycling 90 Unit System) OFF cycle, 100% 18 during ON cycle 110 * High load sliding door opened and closed during testing (A 20-sec door opening event occurring every 15-min) ** Low load doors remained closed during testing The following outlines the necessary steps involved for properly setting up walk-in cooler operation and test runs. Airflow Rate: One of the tasks was to establish a baseline airflow rate inside the walk-in appropriate for its size. Two approaches were used to estimate the necessary airflow rates inside the walk-in cooler due to using ECM. The values obtained from both methodologies were then compared to determine which airflow rate made more sense. The first method was an engineering manual that recommended a minimum of 40 and maximum of 80 air exchanges per hour for holding coolers and freezers (Heatcraft, 2008). Equation 6 shows the relationship between air changes, total airflow rate, and walk-in volume. The manual suggested using the gross interior volume of the walk-in unless the product and equipment occupied more than 10% of the volume. Since the combined volume of the filler products and fan coil (178 ft 3 ) used in this project was about 21% of total internal volume (840 ft 3 ), the net internal volume of (840 ft ft 3 = 662 ft 3 ) was used. This suggested that the airflow rate inside the walk-in cooler can be as low as 441 cfm or as high as 883 cfm Southern California Edison Page 14

29 EQUATION 6. AIRFLOW RATES INSIDE WALK-IN ( cfm 60) Air Changes= Internal Walk -in Volume cfm= OR ( Air Changes Internal Walk - in Volume) 60 The second approach was common industry practices. In this approach, the cooling load of the walk-in cooler was estimated (8,500 to 9,000 Btu/hr) and accordingly an appropriate fan coil size was selected. Reviewing six different fan coils from four different manufacturers (listed below) suggested an airflow rate of 1,400 cfm appropriate for the walk-in cooler size. Russell (low profile): AA26-87B: 2 fans, 6 fins/in, 8,700 Btu/hr at 10 o F TD (+25 o F suction), 1,500 cfm, air defrost Russell (low profile): AA24-84B: 2 fans, 4 fins/in, 8,400 Btu/hr at 10 o F TD (+25 o F suction), 1,660 cfm, air defrost Chandler (low profile): RLC090: 2 fans, fins/in not specified, 9,000 Btu/hr at 10 o F TD (+25 o F suction), 1,400 cfm, air defrost Larkin (low profile): LCA690: 2 fans, 6 fins/in, 9,000 Btu/hr at 10 o F TD (+25 o F suction), 1,400 cfm, air defrost Bohn (low profile): ADT090: 2 fans, fins/in not specified, 9,000 Btu/hr at 10 o F TD (+25 o F suction), 1,400 cfm, air defrost Bohn (Lo-Aire model): LO087: 2 fans, fins/in not specified, 8,700 Btu/hr at 10 o F TD (+25 o F suction), 1,200 cfm, air defrost Comparing these two methodologies revealed that 1,400 cfm seems to be a more realistic and practical airflow rate inside the walk-in. This is because in real-world application a designer or refrigeration contractor would call for an airflow rate of 1,400 cfm based on the cooling load and walk-in box dimensions. Also, to ensure that the discharge air velocity measured at the height level with the fan discharge was not reduced to 25 feet-per-minute (fpm), as suggested by one of the TAG members, 1,400 cfm seemed to be a more reasonable airflow rate. Cooling Load Conditions: As shown in Table 1 the evaporator coil was exposed to two types of load conditions: high and low. A low evaporator load condition refers to the scenario when the walk-in doors remain closed during the entire test period; the total walk-in load is not comprised of infiltration due to door openings. The high evaporator load condition refers to the scenario when a prescribed profile of sliding-door opening events was imposed on the walk-in cooler. Opening the walk-in door allowed to reproduce a realistic scenario by letting the warm and moist air from the adjacent space infiltrate into the refrigerated space. For the number and duration of door opening and closing for the walk-ins, Department of Energy (2010) proposed 60 passages per 24-hour (2.5 passages per hour), 12 seconds per passage, and 15 minutes as the time that door stands open during a 24-hour period (less than 1 minute per hour). Similar approach was used in this project. To create a high load condition for the walk-in, four passages per hour, 20 seconds per passage as the time that door stands open during one-hour period was used. During a 20- sec passage, the door takes 3 sec to open, stays open for 14 sec, and takes 3 sec to close. Southern California Edison Page 15

30 Walk-in cooler temperature and humidity: The refrigeration system was set to maintain a space temperature of +35 o F or saturated evaporator temperature (SET) of +25 o F. In this context, the space temperature refers to the return air temperature (RAT) of the fan coil. The RAT of +35 o F is a common industry practice as well as a requirement by AHRI (2009) when rating the performance of evaporator coils of walk-in coolers. With regards to humidity, although AHRI requires relative humidity level to be less than 50%, no attempt was made to maintain a certain levels of humidity inside the walk-in. The reason for this is that in real-life, there are no humidity controls for walk-ins. 4. Adjacent space temperature and humidity: According to SCE s Foodservice Technology Center, the temperature and humidity in the conditioned kitchen is somewhere between 77 o F and 79 o F with relative humidity of 50%. For an unconditioned kitchen, the temperature can range between 82 o F and 88 o F with up to 65% relative humidity. Accordingly, it seemed reasonable to select and maintain an adjacent space temperature of 80 o F and relative humidity of 60%. This is also in line with engineering models and analysis used for walk-in coolers in foodservice applications (Southern California Edison, 2001). In addition, similar conditions were suggested by one of the TAG members. 5. Dummy or filler products and product simulators: Following ASHRAE (2005) test methods for commercial refrigerators and freezers, the filler product and product simulator types were determined. According to this standard, the product or test simulators should be made of plastic container of at least 1 pint volume with a lid that is filled with sponge material saturated with a solution consisting of a 50/50 ± 2% mixture (by volume) of propylene glycol and distilled water. The temperature shall be measured at the volumetric center when the product simulator height is between 2 in and 2 ¾ in. The shelf spaces that are not occupied by product simulators should be filled with filler or dummy products. The filler products can be wood blocks with an overall density not less than 30 lb/ft 3, and they should occupy between 70% and 90% of the net usable shelf volume. 6. Defrost frequency: Based on intelligence gathered from refrigeration contractors and TAG team, depending on the cooling load of the walk-ins, the defrost frequency can be either four or six defrosts per day. For the purpose of this project, the walk-in refrigeration system s defrost frequency was set to six defrosts per day. So, a four-hour period was used for each test run. A four-hour test period seemed to be sufficient since the defrost efficiency was not part of this study. Task 5: Fabricate CO 2 tracer gas probes This task fabricates the CO 2 probe structure and layout within the TTC s walk-in test unit. This was accomplished prior to conducting the infiltration analysis testing. Task 6: Measure infiltration rate using CO 2 tracer gas technique This task maps the infiltration rate as a function of the duration of the door opening. Below summarizes three sub-tasks for achieving the main goal. This was accomplished by first measuring the infiltration rate with completely closed door (sub-task 6.1) and completely open door (sub-task 6.2).The data gathered set the upper and lower limits of infiltration rate of walk-in. In addition, infiltration rates were measured with several prescribed frequency and duration of door opening (sub-task 6.3). Task 6.1: Measure infiltration rate with completely closed door Task 6.2: Measure infiltration rate with completely open door Task 6.3: Measure infiltration rate with several prescribed duration of door opening times. 3 sec, 10 sec, 30 sec, 120 sec, 300 sec, 400 sec, and 600 sec are examples of door opening times that were explored. Not all of these times were explored for all permutations: Since infiltration decays at different rates dependent upon other varied driving conditions (fan speed, Southern California Edison Page 16

31 Task 7: ambient/walk in temperatures), the door opening times were chosen to give the best resolution, based upon the upper and lower limits (tasks 6.1 and 6.2). Identify the impact of varying fan speed on infiltration rate This task quantifies the impact of fan speed on the infiltration rate. This was accomplished by measuring infiltration rate under three different fan speeds: low (50%, 725 cfm), intermediate (75%, 1,090 cfm), and maximum (100%, 1,400 cfm) speed. Task 8: Identify the impact of temperature variations between the walk-in and the adjacent space on infiltration rate This task quantifies the impact of temperature gradients between the walk-in and adjacent space on the infiltration rate. In addition, this task establishes the order of magnitude effect of temperature gradients on infiltration rate. This was accomplished by using the following four adjacent space conditions: o F dry-bulb / 60% relative humidity (representing conditioned kitchen) o F dry-bulb / 55% relative humidity (representing conditioned space) o F dry-bulb / 14% relative humidity (representing hot and dry condition) o F dry-bulb / 82% relative humidity (representing hot and humid condition) Task 9: Identify the impact of door opening duration, strip curtains, propped open door/inadequate door gasket seals, door types (swing vs. sliding), and door open area on infiltration rate This task quantifies the impact of infiltration barriers like strip curtains (sub-task 9.1) as well as propped open door/inadequate door gasket seal (sub-task 9.2) on infiltration rate. To simulate a poor gasket seal, the swinging door was propped open. The gasket continued to touch the perimeter of the doorway, but was not pressing down to complete a seal. Also, the goal was to determine if there are any differences in infiltration rate as a function of door types, specifically swing type versus sliding doors (sub-task 9.3). Additionally, the infiltration rate was quantified as function of door open area (sub-task 9.4). Task 9.1: Measure infiltration rate with strip curtain and completely open door Task 9.2: Measure infiltration rate with propped open door Task 9.3: Measure infiltration rate with open swing type door Task 9.4: Measure infiltration rate as a function of door open area at three different increments of width (25%, 50%, and 75% of door width) Task 10: Map the velocity profile of infiltrated warm air and exfiltrated cold air This task maps the velocity profiles due to infiltration or discharge process at the doorway. This was accomplished by using flow visualization technique, namely particle image velocimetry. Task 11: Data Analysis This task reviews and analyzes the collected data. This includes numerical and graphical presentation of data. Task 12: Prepare the final report This task develops and finalizes the project report. Southern California Edison Page 17

32 WALK-IN COOLER DESCRIPTION Figure 1 shows the walk-in cooler s layout and key dimensions as well as the target dry-bulb (DB) temperature and relative humidity (RH) inside and outside of the cooler. The walk-in panel insulation was 4-inch rigid polyurethane with white stucco galvanized exterior and interior finish with an overall thermal resistance value (R-value) of 25 ft 2 - o F-hr/Btu. The walk-in had one horizontal sliding door and one standard hinged door. The motorized sliding door can be set to automatically open and close, and the time that it takes to open or close the sliding door can be controlled and adjusted. Standard Hinged Door 3' x 7' Walk-in Cooler 9'-4"L x 9'-4"W x 9'-8'H 35 o F DB Sliding Door 3' x 7' Adjacent Space 80 o F DB / 60% RH NOT Running 70" 20" 48" Running Adjacent Space 80 o F DB / 60% RH 1' Evaporator Coil 22" (run one fan) 6'-4" Room 2 Room 4 20" Adjacent Space 80 o F DB / 60% RH 2' Walk-in Cooler Layout SCALE: NONE FIGURE 1. LAYOUT OF THE WALK-IN COOLER USED The walk-in was equipped with a single fan coil that had two refrigeration circuits and two half-hp evaporator fans (Figure 2). The fan coil was an A series unit cooler, model number A from Century Refrigeration. The factory installed fan motors were three-phase 460V induction motors. The fan coil or unit cooler can provide airflow rates of up to 7,860 ft 3 /min (cfm) or 740 ft/min (fpm). For the purpose of this project, one of these fans was replaced by a single-phase 120V, quarter-hp ECM connected to a VSD. So, the required velocity or airflow in the walk-in cooler was re-evaluated to ensure the ECM fan motor speed was set properly. Southern California Edison Page 18

33 FIGURE 2. PHOTOGRAPH OF EVAPORATOR FAN COIL OF WALK-IN COOLER The fan coil is served by a multiplex compressor rack system (Figure 3) that has four Bitzer octagon reciprocating compressors (two 4FC-3.2, one 2FC-3.2, and one 2CC-4.2). The compressor rack is capable of matching capacity with refrigeration load either by using unloaders and/or modulating the compressor speed. R-404A refrigerant is used. FIGURE 3. PHOTOGRAPH OF REFRIGERATION COMPRESSORS The DB temperature and RH in the spaces adjacent to the walk-in cooler were controlled by independent air-handling units (AHUs). These AHUs were served by a separate multiplex refrigeration rack system, independent of the test walk-in cooler refrigeration rack. Southern California Edison Page 19

34 TEST METHODOLOGY This laboratory assessment project was conducted at SCE s TTC (see Appendix F). The assessment was comprised of two studies. One study focused on addressing infiltration rate in walk-ins. The other study focused on addressing feasibility of reducing fan speed in walkins. INFILTRATION RATE STUDY Series of tests were performed to study the infiltration rate and also visualize the flow pattern at the door during this process. CO 2, an inexpensive and safe tracer gas, was used to monitor the infiltration rate. The process consisted of bringing the tracer gas concentration to a stable level inside the cooler when the door was closed and monitor the concentration during the opening period. After the closing of the door, the concentration in the room became uniform due to the operation of the fan and the convective motion of air. The difference between the initial and final concentration at steady state conditions was correlated to the amount of CO 2 loss during infiltration. This concentration was related to the amount of cold air that exfiltrates out of the cooler or the warm air that infiltrates into the cooler. The derivation of these equations is given in Appendix A. TEST DESIGN AND INSTRUMENTATION The experimental set up and the gas analyzer equipment are shown in Figure 4. Several probes, as shown in Figure 5, were installed throughout the cooler to collect samples and obtain a good representation of the average CO 2 concentration inside the cooler. To accurately measure the average tracer gas concentration in the cooler, it was necessary to position the sampling probes properly to collect the non-uniform samples in the room. This non-uniformity was caused by air motion induced by the fan and some possible slow moving air pockets that came into thermal equilibrium with the rest of the air inside the room after a longer elapsed time during or after infiltration. The CFD analysis was performed to determine those regions that here are referred to as pseudo dead pockets of air inside the cooler. The results are detailed in Appendix D. Overall, an initial uniform cooler temperature of 25 C or 77 F was assumed. A cold evaporator temperature was set and the fan started blowing cold air inside the cooler. The temperature inside the cooler at numerous spots was monitored and the variation of temperature in these spots was observed. Those regions that came into thermal equilibrium with the rest of the cooler during the cool down process were identified as pseudo dead pockets and sampling probes for the tracer gas were installed in those regions, or close to them, for proper sampling. Furthermore, the tracer gas concentration in the cooler was monitored after the door closing until a uniform concentration in the room was observed. These collected samples were taken to the gas analyzer for data login. Suction pumps drew the mixture of tracer gas and air from the desired points and transferred it to the gas analyzer. The pumps were installed between the sampling probes and the gas analyzer and each gas analyzer channel required its own dedicated pump. The outside area where the warm air was infiltrating into the cooler Southern California Edison Page 20

35 was a controlled environment room, and therefore could not act like a mass sink. To retain accuracy of the data, the CO 2 concentration was monitored inside this room. The maximum amount was only about 6% of the total amount of the initial tracer gas, and was accounted for throughout the calculations. FIGURE 4. A) CO 2 TRACER GAS TANKS (LEFT PHOTO) B) GAS ANALYZER, HORIBA VA-3000, (RIGHT PHOTO) This gas analyzer used a non-dispersive infrared (NDIR) technology for measuring the concentration of the CO 2 tracer gas. This instrument included three input channels, as there were three sample inlets to the analyzer. The rate at which the samples were analyzed and concentrations were reported was one data per second. By using Horiba s DL-3000 software, the data was logged and stored in a computer. The repeatability, linearity and zero/span shift errors of the gas analyzer were 0.5%, 1%, and 0.5% of the full scale value, respectively; while the sensitivity of the instrument was less than 1%. The resolution of the instrument was 1 parts-permillion (ppm). The overall uncertainty was about ± 480 ppm equivalent to ±1.9% of the full scale value. For calibrating the gas analyzer, the system was first zeroed by a gas with tracer gas concentration equal to zero. Next, a span gas with a known concentration of the tracer gas was flown into the gas analyzer. By performing these two steps the analyzer recognized its linear calibration through these two concentrations. In these experiments nitrogen was the zero gas and the span gas contained 22,450 ppm CO 2 (equivalent to 2.245% concentration). The flow rate of the calibration gases and the experimental samples was 0.5 liter per min. Southern California Edison Page 21

36 FIGURE 5. LOCATION OF PROBES CORRESPONDING TO SECOND AND THIRD GAS ANALYZER CHANNELS The PIV equipment consisted of a 200 millijoules (mj) per pulse, dual-head, pulsed neodymium-doped yttrium aluminium garnet, also known as Nd:YAG, laser. This laser with five nanosecond pulse duration was used to instantaneously illuminate approximately a 127 millimeter (mm) or five-inch square field of view in the flow. The laser light was passed through suitable optics to form a light sheet with approximately one mm thickness. The flow was seeded with micrometer (μm) particles consisting of a mixture of glycerin and water (smoke). The laser light scattered from the seeded particles was imaged on a camera with a 1,008 by 1,018 pixel charged-couple device array. The images were processed with PIV PROCessing code, which is a correlation-based processing software developed by Wernet (1999) that uses sub-region shifting and multi-pass correlation. A 64 by 64 pixel region was used on the first pass and a 32 by 32 pixel region with 50% overlap was used on the second pass. Figure 6 shows the setup for the visualization experiment via PIV. Southern California Edison Page 22

37 FIGURE 6. A) THE CAMERA MOUNTS FOR THE SLIDING DOOR (LEFT PHOTO) B) THE LASER AND MIRROR MOUNT (RIGHT PHOTO) VARYING EVAPORATOR FAN SPEED STUDY A series of tests were performed to study the effects of varying the speed of the evaporator fan on total power demand and energy usage of a refrigeration system and its components, excluding the condenser. Additionally, effects of varying the speed of the evaporator fan on product temperature and temperature stratification inside the walk-in was captured. TEST DESIGN AND INSTRUMENTATION Since there were no established standards for testing the performance of walk-in coolers and freezers, the performance testing relied primarily on best engineering practices. Whenever applicable, however, AHRI (2009) Standard 1250 and ASHRAE (2005) Standard 72 were used. For example, the measurements for DB temperature and RH of air entering the evaporator coil followed AHRI (2009) Standard 1250, whereas type of filler products and product simulators followed ASHRAE (2005) Standard 72. The sliding door was set to automatically open and close for a specified period of time to impose high and moderate cooling load conditions. The operation of compressor and evaporator fan VSD was controlled based on RAT of 35 ± 2 F, for all test scenarios. Specifically, when the RAT reached 33 F, the refrigeration system stopped and the speed of evaporator fan motor was dropped from 100% to 70% of full speed. A call for cooling was initiated and the evaporator fan motor speed was set back to 100% of full speed when the RAT reached 37 F. Additionally, to duplicate a practical scenario, the temperature and humidity in the spaces adjacent to the walk-in cooler where the testing was performed were controlled by independent AHUs. Southern California Edison Page 23

38 The National Instruments SCXI data acquisition system was used to log the test data. The data acquisition system was set up to scan 186 data channels in 20-second intervals and log data in one-minute 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 TTC 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. 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 a customized refrigeration analysis model where detailed calculations were performed. All 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 Southern California Edison Page 24

39 Table 2 provides the specifications and calibration dates for all the sensors used in this project. The following lists the core monitoring points. power and runtime evaporator fan compressor air side DB and RH inside the walk-in DB and RH outside of the walk-in, adjacent space DB and RH of evaporator inlet air, RAT DB and RH of evaporator outlet air, DAT air velocities at the evaporator inlet air velocities at evaporator outlet or discharge refrigerant side refrigerant temperature and pressure at the inlet of the evaporator refrigerant temperature and pressure at the outlet of the evaporator refrigerant temperature and pressure at the inlet of the compressor refrigerant temperature and pressure at the outlet of the compressor refrigerant mass flow rate other measurements evaporator fan speed barometric pressure inside the walk-in pressure differential between walk-in and adjacent space walk-in panels (all six sides) interior and exterior temperatures condensate mass Southern California Edison Page 25

40 TABLE 2. SPECIFICATIONS AND CALIBRATION DATES FOR SENSORS USED SENSOR TYPE Temperature (type-t thermocouples) Temperature (resistance temperature device, RTD) Humidity Dew Point Velometer Pressure (0-500 psi) Pressure (0-100 psi) Pressure (0-10 inches of water, in-wg) MAKE/MODEL Kaye Instruments, T/W 50 through 80, melt # 8032 Hy-Cal Engineering, RTS-37-A-100 Vaisala, HMP233 Edgetech, Dew Prime DF Dew Point Hygrometer TSI Inc., 8455 Air Velocity Transducer Ashcroft, AQS Ashcroft, AQS Ashcroft, AQS ACCURACY (NIST TRACEABLE) ± 0.01 C (± F) ± 0.1 C (± 0.18 F) ± 0.2 C (± 0.36 F) ± 0.2 C (± 0.36 F) ± 0.5% of full scale, ± 2.0% of reading ± 0.025% of full scale ± 0.025% of full scale ± 0.06% of full scale CALIBRATION DATE CALIBRATION DUE DATE Power Fluke, 435 ± 1% of reading Refrigerant Mass Flow Meter Rotational Speed Scale Micro Motion, (Coriolis meter) F025SR113SU Monarch Instrument, ACT-3 HP-30K ± 0.1% of rate ± 0.001% of full scale ± 0.1 gram (± ounces) Figure 7 illustrates the quantity and location of discharge or supply air DB and dew point (DP) temperature sensors. As shown, air DB temperatures were measured at five locations. For DP, sampling tubes were arranged to draw the air from four different locations. The sampling points or locations were in close vicinity of the DB temperature measurements. Figure 7 also shows that the evaporator fan motor power and the rotational speed (in terms of revolutions per minute (rpm)) were measured. Figure 8 illustrates the location of discharge air velocities. As shown, the discharge air velocities were measured at three different locations at the height level with the fan. Southern California Edison Page 26

41 Ceiling 1" 6'-4" 13" 20" 5" 5" TDB1 TDP1 TDP2 TDB2 Sampling Tube (TDP) 25" Adjacent Wall TDB3 Adjacent Wall 2' 5" 5" TDB4 TDP3 TDP4 TDB5 Fan kw Fan RPM NOT Running 1' 5" 5" 5" 5" 48" 64" Legend: Fan kw = KW of the evaporator fan (1 sensor) Fan RPM = RPM of the evaporator fan (1 sensor) TDB = Air dry-bulb temperature (5 sensors) TDP = Air Dew-point temperature (4 samples) Evaporator Coil Front View SCALE: NONE FIGURE 7. QUANTITY AND LOCATION OF SENSORS FOR MEASURING EVAPORATOR DISCHARGE OR SUPPLY AIR DRY-BULB AND DEW-POINT TEMPERATURES, EVAPORATOR FAN MOTOR POWER AND ROTATIONAL SPEED Ceiling 1" 70" 20" DAV3 DAV2 DAV1 10" 60" 40" 20" Evaporator Coil 22" 25" DAV3 0 DAV1 0 Top View DAV2 0 Adjacent Wall Adjacent Wall Sliding Door 7' Legend: DAV = Discharge air velocity (3 sensors) 90" 3' Floor Discharge Air Velocity SCALE: NONE FIGURE 8. QUANTITY AND LOCATION OF SENSORS FOR MEASURING DISCHARGE OR SUPPLY AIR VELOCITY Southern California Edison Page 27

42 Figure 9 shows the quantity and location of return air DB and DP temperature sensors as well as the air velocity sensors. Similar to discharge air, return air DB temperatures were measured at five locations, and DP sampling tubes were arranged to draw the air from four different locations. The air velocities were measured at four locations. Figure 9 also illustrates the sensor locations for measuring refrigerant temperature and pressure at the inlet and outlet of the evaporator coil. Ceiling 1" TDB1 Sampling Tube (TDP) TDB2 Adjacent Wall NOT Running V1 TDP1 TDP2 V2 6" 6" 13" Adjacent Wall Flow 1' 24" TXV Flow Tref Pref TDP3 TDP4 6" Tref Pref Dpipe Evaporator Outlet to Compressor Suction Line TDB4 V3 TDB3 Sampling Tube (TDP) 7.5" 7.5" 7.5" 7.5" 30" V4 TDB5 6" Within 6 x Dpipe Evaporator Inlet (Liquid Line) 2' 64" 48" Legend: Pref = Refrigerant pressure (2 sensors; 1 at inlet, 1 at outlet) TDB = Air dry-bulb temperature (5 sensors) TDP = Air Dew-point temperature (4 samples) Tref = Refrigerant temperature (2 sensors; 1 at inlet, 1 at outlet) V = Air velocity (4 sensors) Evaporator Coil -- Rear View SCALE: NONE FIGURE 9. QUANTITY AND LOCATION OF SENSORS FOR MEASURING EVAPORATOR RETURN AIR DRY-BULB AND DEW-POINT TEMPERATURES, EVAPORATOR RETURN AIR VELOCITY, REFRIGERANT TEMPERATURE AND PRESSURE ENTERING AND LEAVING THE EVAPORATOR COIL Figure 10 depicts the location of DB temperature and RH sensors used to measure the walk-in s interior and exterior (adjacent space) air temperature and humidity. Inside the walk-in, two poles were set up, one close to the sliding door and one distant from the sliding door (Figure 11). On each pole, ten temperature sensors at one-foot increments and one RH sensor at the six-foot level were installed. This captured the air temperature stratification along the cooler height at two locations. Similar set up was used to measure temperature and humidity in the adjacent spaces, and in front of the sliding door. The walk-in s interior and exterior surface temperatures were monitored as well. As shown in Figure 10 and Figure 12, for each side of the cooler, two sensors were used to measure the interior and two sensors to measure the exterior surface temperatures. Figure 10 also depicts the quantity and location of product simulators used in this project. Figure 12 shows a photograph of a product simulator. A total of four racks were used to capture the product temperatures along each side of the walk-in. Each rack had four shelves, and on each shelf three product simulators were used. Overall, 48 product simulators were used to measure product temperatures. Southern California Edison Page 28

43 TWall Standard Hinged Door 3' x 7' Temperature & Humidity Pole (10 TDB and 1 RH) Adjacent Space TWall Product Shelf 60" L x 24" W x 75" H [5' L x 2' W x 6.25' H] 4-shelf Product Shelf 46" 26" 22" 60" L x 18" W x 75" H [5' L x 1.5' W x 6.25' H] 4-shelf 49" 26" Temperature & Humidity Pole 66" (10 TDB and 1 RH) [5.5'] 60" [5'] Walk-in Cooler 9'-4"L x 9'-4"W x 9'-8'H +35 o F DB Evaporator Coil (run one fan) 6'-4" Product Shelf 60" L x 24" W x 75" H [5' L x 2' W x 6.25' H] 4-shelf 24" TWall 24" Temperature Pole (10 TDB) Sliding Door 3' x 7' Temperature & Humidity Pole (10 TDB and 1 RH) Adjacent Space 30" 30" PT1 PT2 PT3 Shelf 24" PT Center PT PT PT4 PT5 PT6 Product Shelf 60" L x 24" W x 75" H [5' L x 2' W x 6.25' H] 4-shelf PT7 PT8 21" PT9 21" Sliding Door Will Remain Open During Testing for Proper Mixing of Surrounding Air TDB TWall Adjacent Space TDB PT10 PT11 PT12 Product Temperature Locations on the Shelves Legend: PT = Product Temperature (48 sensors) RH = Relative humidity (4 sensors) TDB = Air dry-bulb temperature (52 sensors) TWall = Wall surface temperatures (sides, ceiling and floor; total of 22 sensors) Standard Hinged Door Will Remain Open During Testing for Proper Mixing of Surrounding Air Wall Surface Temperature, Temperature/Humidity Poles, and Dummy Product Locations SCALE: NONE FIGURE 10. QUANTITY AND LOCATION OF SENSORS FOR MEASURING WALK-IN S INTERIOR AND EXTERIOR DRY- BULB TEMPERATURES AND RELATIVE HUMIDITIES, INTERIOR AND EXTERIOR WALK-IN PANEL SURFACE TEMPERATURES, AND PRODUCT TEMPERATURES Temperature/humidity pole away from sliding door Temperature/humidity pole close to sliding door Filler Products FIGURE 11. PHOTOGRAPH OF FILLER PRODUCTS AND TEMPERATURE POLES INSIDE THE WALK-IN FIGURE 12. A) PHOTOGRAPH OF WALL SURFACE TEMPERATURE SENSORS (LEFT PHOTO) B) PHOTOGRAPH OF A PRODUCT SIMULATOR (RIGHT PHOTO) Southern California Edison Page 29

44 DATA 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 XProps TM refrigerant property program, version 1.5. XProps TM and was also used to determine the saturated refrigerant temperatures based on collected temperature and pressure data. 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 7). EQUATION 7. REFRIGERATION EFFECT RE = h evap-out h evap-in where, RE = Refrigeration effect of the refrigerant in the evaporator, (Btu/lb) h evap-out h evap-in = Superheated refrigerant enthalpy at the evaporator exit, (Btu/lb) = Sub-cooled liquid refrigerant enthalpy at expansion valve inlet, (Btu/lb) The refrigeration load of the case is the rate of cooling or heat removal (in BTU) that takes place at the evaporator of the walk-in per hour (Equation 8). This quantity is obtained by multiplying the refrigeration effect by refrigerant mass flow rate, which is extracted from the data acquisition system. The total cooling load was determined by using Equation 8. Southern California Edison Page 30

45 EQUATION 8. TOTAL REFRIGERATION LOAD (IN BTU/HR) Q Total - refrig = m refrig RE k where, Q Total - refrig = Total refrigeration load of the case, (Btu/hr) m refrig = Mass flow rate of refrigerant, (lb/min) RE k = Refrigeration effect of the refrigerant in the evaporator, (Btu/lb) = Conversion factor, (60 min/hr) To determine the refrigeration load of the walk-in in cooling tons, it can be divided by 12,000, a conversion factor for Btu/hr to tons (Equation 9). EQUATION 9. TOTAL REFRIGERATION LOAD (IN COOLING TONS) QTotal- refrig Q Total- refrig (tons) = 12,000 where, Q Total - refrig(tons) = Total refrigeration load of the case, (tons) Q Total - refrig = Total refrigeration load of the case, (Btu/hr) The volume flow rate of air circulated throughout the walk-in is obtained using Equation 10. This parameter was obtained using the average air velocity and opening area at the inlet of the evaporator coil. EQUATION 10. VOLUMETRIC AIR FLOW RATE cfm air = V avg-inlet-air A coil-inlet where, cfm air = Volumetric flow rate of air in the walk-in, (ft 3 /min) V avg-inlet-air = Average velocity of the air at the inlet of the evaporator coil, (ft/min) A coil-inlet = Total inlet area of the evaporator coil, (ft 2 ) One of the system parameters is the evaporator coil superheat. The evaporator coil superheat was determined 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 evaporator outlet according to Equation 11. Southern California Edison Page 31

46 EQUATION 11. EVAPORATOR COIL SUPERHEAT SH evap = T v SET where, SH evap = evaporator coil superheat, F T v = vapor refrigerant temperature at the outlet of the evaporator coil, F SET = saturated evaporating temperature based on evaporator outlet pressure, F Equation 12 was used to determine the condenser sub-cooling. Condenser sub-cooling was obtained by subtracting the liquid refrigerant temperature at the outlet of the condenser from the saturated condensing temperature based on compressor outlet pressure. EQUATION 12. CONDENSER COIL SUB-COOLING SC cond = SCT T L where, SC cond = condenser coil sub-cooling, F SCT = saturated condensing temperature based on condenser outlet pressure, F T L = liquid refrigerant temperature at the outlet of the condenser coil, F The total system power and energy use for the tests excluded condenser power. The total system power of the fixture was obtained using Equation 13. The power usage associated with the evaporator fan motors and compressor was read directly from the data acquisition system. The lighting power and auxiliary heaters were based on the rated wattage. Auxiliary heaters were used for the multiplex compressor test scenarios only. EQUATION 13. TOTAL REFRIGERATION POWER USAGE, EXCLUDING CONDENSER kw Total = kw evap-fan + kw light + kw aux-heaters + kw comp where, kw Total kw evap-fan kw light kw aux-heaters kw comp = power usage by the refrigeration system, excluding condenser, kw = power usage by the evaporator fan motors, kw = power usage by the light in the walk-in, kw = power usage by the auxiliary heaters in the walk-in, kw = power usage by the refrigeration compressor, kw The energy consumption of the evaporator fan motors, lights, auxiliary heaters, and the compressor is defined as the product of supplied power and total hours of power usage. Equation 14 shows the general format for obtaining energy usage. Once the energy usage of each component was determined, the total was obtained by adding all the individual components together. Southern California Edison Page 32

47 EQUATION 14. ENERGY USAGE where, kwh = energy usage by the end-use, kwh kw(t) = power usage by the end-use, as a function of time, kw t = time of power usage by the end-use, hours The conduction load refers to the conduction of heat through the walk-in panels. The temperature difference between the interior and exterior surface of the walk-in panels is the driving force for this transfer of heat. The conduction also is dependent on the overall coefficient of heat transfer (also known as U-value) and the area of the walk-in panels. The U-value, however, is the reciprocal of R-value. As mentioned previously, the R-value of the walk-in panels were 25 ft 2 - o F-hr/Btu. Equation 15 was used to obtain the overall conduction load by adding the conduction load through each side of the walk-in cooler. EQUATION 15. CONDUCTION OR TRANSMISSION LOAD OF THE WALK-IN Q é conduction = å êu ë panel Apanel Texterior- surface - T int erior- surface where, ( ) Q conduction = Conduction or Transmission load of the walk-in, Btu/hr U panel = Overall coefficient of heat transfer for the panels, 1/R, (Btu/hr-ft 2 - F) A panel = Surface area of the walk-in panel, ft 2 T exterior-surface = Exterior surface temperature of panel, F T interior-surface = Interior surface temperature of panel, F ù úû Southern California Edison Page 33

48 RESULTS RESULTS OF INFILTRATION RATE STUDY QUANTIFICATION OF INFILTRATION RATES USING TRACER GAS METHOD Each infiltration rate plot/figure is a compilation of the average infiltration rates, determined using Equation 4, for various door opening times (3 sec, 10 sec, 30 sec, etc.) and different fan speeds (100%, 75%, 50%, OFF) at fixed ambient space conditions. Different figures present results at various ambient space conditions. Figure 13 shows the average infiltration rates as a function of fan speed and the time that the door stayed open when the adjacent condition was 75 F DB and 55% RH. It shows that the initial infiltration rate started at around 39 ft 3 /sec, and the significant portion of the infiltration process continued for about 122 seconds. It should be noted that as the infiltration rate decays, the volume of cold air left in the cooler also reduces. The amount of cold air left inside the cooler can be calculated (represented) by the area under the infiltration rate curve. The infiltration rate asymptotically goes to nearly zero as the period of time that the door is left open increases in all test cases. For instance, it takes approximately 5 minutes of door opening to effectively have the entire cold air in the cooler to be replaced by outside air. This concept has been mathematically explained in Appendix E. 75% of the infiltration occurred during the first 30 seconds from the time that the door was opened and left opened. In other words, 75% of the cold air inside the walk-in was exfiltrated out of the cooler during the first 30 seconds of the door s stay open time. Calculating the area under the curve or total infiltrated volume, it turned out that it was about 640 ft 3, which was very close to the total empty volume of the walk-in cooler. This was actually a verification of the method that indicated the decaying nature of the infiltration for the cooler as the inside and outside air came into equilibrium. In addition, Figure 13 shows that the fan speed did not seem to play a major role in the infiltration process. In effect, only 5% of the infiltration difference could be attributed to the fan speed, which was within the experimental errors. Figure 14 depicts the infiltration rate as a function of fan speed and the time that the door stayed open when the adjacent condition was 80 F DB and 60% RH. Comparing Figure 13 and Figure 14 revealed that increased temperature difference between the cold walk-in and warm adjacent space causes initial infiltration rates to increase and causes the infiltration process to occur faster. The initial (maximum) infiltration started at around 41 ft 3 /sec and the infiltration process effectively lasted for about 105 seconds. In addition, Figure 14 shows that the fan speed did not seem to play a major role in the infiltration process. Figure 15 shows the infiltration rate as a function of the time that the door stayed open and fan speed when the adjacent condition was 84 F DB and 82% RH. Comparing Figure 14 and Figure 15 showed that the total infiltration in the 600 sec time period has decreased. This decrease is attributed to increased humidity of the air in the adjacent space (82% RH vs. 60% RH), although the DB temperature was Southern California Edison Page 34

49 Infiltration Rate (ft³/s) Investigating the Effects of Fan Speed Variation and slightly higher (84 F vs. 80 F). That is, the humidity of the adjacent space has had an adverse effect on the infiltration. Additionally, the fan speed did not seem to play a major role in the infiltration process (Figure 15). Figure 16 shows the infiltration rate as a function of the time that the door stayed open and fan speed when the adjacent condition was 115 F DB and 14% RH. As shown, the increased air temperature in the adjacent space did increase the infiltration rate. With the initial infiltration rate of about 65 ft 3 /sec, it was the highest initial infiltration rate compared to the other three adjacent space conditions. This was mainly due to the increased density gradient. Another interesting result was found when the fan was set off during the infiltration. The tracer gas method predicted about 20% less infiltration for the fan off condition during the first 15 seconds of the process. Therefore, the impact of the fan speed during the infiltration process was mainly a function of temperature difference between the warm adjacent space and cold refrigerated space Tracer Gas Method - Fan Off 60 Tracer Gas Method - 50% Fan Tracer Gas Method - 75% Fan 50 Tracer Gas Method - 100% Fan Time (sec.) *Infiltration rates are based on methods using measured original cold air exfiltration FIGURE 13. INFILTRATION RATE AS A FUNCTION OF THE TIME DOOR STANDS OPEN AND FAN SPEED FOR 75 F DRY-BULB AND 55% RELATIVE HUMIDITY TRACER GAS METHOD Southern California Edison Page 35

50 Infiltration Rate (ft³/s) Infiltration Rate (ft³/s) Investigating the Effects of Fan Speed Variation and Tracer Gas Method - Fan Off Tracer Gas Method - 50% Fan Tracer Gas Method - 75% Fan Tracer Gas Method - 100% Fan Time (sec.) FIGURE 14. INFILTRATION RATE AS A FUNCTION OF THE TIME DOOR STANDS OPEN AND FAN SPEED FOR 80 F DRY-BULB AND 60% RELATIVE HUMIDITY TRACER GAS METHOD Tracer Gas Method - Fan Off Tracer Gas Method - 50% Fan Tracer Gas Method - 75% Fan Tracer Gas Method - 100% Fan Time (sec.) FIGURE 15. INFILTRATION RATE AS A FUNCTION OF THE TIME DOOR STANDS OPEN AND FAN SPEED FOR 84 F DRY- BULB AND 82% RELATIVE HUMIDITY TRACER GAS METHOD Southern California Edison Page 36

51 Infiltration Rate (ft³/s) Investigating the Effects of Fan Speed Variation and Tracer Gas Method - Fan Off Tracer Gas Method - 50% Fan Tracer Gas Method - 75% Fan Tracer Gas Method - 100% Fan Time (sec.) *Infiltration rates are based on methods using measured original cold air exfiltration FIGURE 16. INFILTRATION RATE AS A FUNCTION OF THE TIME DOOR STANDS OPEN AND FAN SPEED FOR 115 F DRY-BULB AND 14% RELATIVE HUMIDITY TRACER GAS METHOD COMPARISON OF INFILTRATION RATES (TRACER GAS VS. COMMON METHODS) This section compares the estimated infiltration rates using Tamm s, Gosney s and Cleland s methods (plots are included in Appendix E) with results obtained from the tracer gas method. Comparisons were made for four different adjacent space DB and RH conditions under four different fan speed scenarios. The comparison revealed that none of the previous models or methods did account for changes in densities inside and outside of the walk-in cooler as a function of time. In addition, they did not take into account the net air volume inside the walk-in cooler. The net air volume refers to the total interior volume of the walk-in minus the volume that is occupied by the food products and equipment like evaporator coil inside the walk-in. So, since the air volume inside the walk-in is predetermined and finite, any changes in the net volume will change the available air volume to be displaced or exchanged between adjacent warm space and refrigerated space. As a result, since the earlier models do not consider a finite volume for the cooler, therefore, they were not curate for the entire time period in which the infiltration process took place. Overall, since these earlier models were not developed for a finite volume space and density variations, they will fail after a certain threshold. Both Tamm s and Gosney s model under-predicted the infiltration rates. However, for total infiltration volume, it seemed that both Tamm s and Gosney s model predictions may be correct for the first seconds of infiltration process, but was invalid for longer times. Cleland s model over-predicted the initial infiltration significantly and therefore was only valid for a shorter period of time, about seconds. Cleland used an orifice analogy to use the standard flow meter equation with the calibration factor. After modifying the original model for abrupt expansion effect, the over-prediction became about 35% Southern California Edison Page 37

52 less than the original equation. In Cleland s model the effect of fan speed was embedded in the pressure difference. Tamm and Gosney did not include this effect. NATURAL INFILTRATION RATE (NATURAL DECAY) Figure 17 shows the loss of adjacent space air in the walk-in cooler due to natural infiltration. The decay of the adjacent space air in the walk-in cooler was directly proportional to the decay of the CO 2 concentration. The embedded figure (left bottom corner of Figure 17) is a magnification of the earlier stages. It appeared that the natural infiltration was linear in time and it was also a function of fan speed, temperature and humidity differentials between the outside and inside air. The data has been extrapolated beyond the original data points (see Appendix B). The natural infiltration took place in a longer timeframe as compared to the infiltration that occurred after the door was opened and kept open for about seconds. This signified the importance of well-sealed and insulated system. FIGURE 17. THE NATURAL DECAY IN THE COOLER (CLOSED DOOR AND STABILIZED SYSTEM) AS A FUNCTION OF FAN SPEED AND ADJACENT SPACE DRY-BULB TEMPERATURE AND RELATIVE HUMIDITY Southern California Edison Page 38

53 IMPACT OF DOOR TYPE (SWING VS. SLIDING) ON INFILTRATION Several experiments were performed to measure the infiltration through swinging or hinged type door and the results were compared to the infiltration obtained using sliding type door (Figure 18). The results revealed that the infiltration were consistent for the first 40 seconds and started to differentiate afterwards. This was attributed to the layout of the room, resulting in a more corridor-like entrance before the door. There was less abrupt expansion when compared to the environmentally controlled room off the sliding door. Also, this was partially attributed to fan being offset farther away from the sliding door. Nonetheless, the infiltration through the swinging door was very similar to those conducted earlier. FIGURE 18. INFILTRATION COMPARISON FOR SLIDING AND SWING TYPE DOORS AT THE SAME CONDITIONS (FULL SPEED FAN AND ADJACENT SPACE CONDITION OF 75 F DRY-BULB AND 55% RELATIVE HUMIDITY) IMPACT OF STRIP CURTAINS AND PROPPED OPEN DOOR/INADEQUATE DOOR GASKET SEAL ON INFILTRATION The impact of strip curtains and a propped open door/inadequate door gasket seals on the infiltration were also captured. As shown in Figure 19, the natural decay was seen as a linear trend when the cooler remained closed for a significant duration of time. To simulate a poor gasket seal, the swinging door was propped open. The gasket continued to touch the perimeter of the doorway, but was not pressing down to complete a seal. As shown, infiltration increased for the poor gasket condition when compared to the natural decay profile. For strip curtain test scenario, after a vinyl strip curtain was installed, the sliding door was left open and undisturbed during the test duration. Comparing infiltration with and without the strip curtain, showed that the strip curtains can minimize the infiltration, as the door remains open. Southern California Edison Page 39

54 FIGURE 19. INFILTRATION COMPARISON OF STRIP CURTAIN AND PROPPED OPEN DOOR/INADEQUATE DOOR GASKET SEAL (ADJACENT SPACE CONDITION OF 75 F DRY-BULB AND 55% RELATIVE HUMIDITY, AND FULL FAN SPEED) IMPACT OF STRIP CURTAINS AND DOOR PASSAGES ON INFILTRATION The complete infiltration data was also plotted on the same graph to show the upper bound of the problem. In-and-out traffic was introduced to the strip curtain with the sliding door being open for a 30-second time frame. Strip curtains (ones that are free of defects and properly installed) did significantly reduce infiltration of outside air into the cooler as seen on Figure 20. Southern California Edison Page 40

55 FIGURE 20. INFILTRATION COMPARISON BETWEEN NO BARRIER AND STRIP CURTAIN WITH IN/OUT TRAFFIC (ADJACENT SPACE CONDITION OF 75 F DRY-BULB AND 55% RELATIVE HUMIDITY, AND FULL FAN SPEED) IMPACT OF DOOR OPENING AREA ON INFILTRATION Since the sliding door could be manually left in any position on its track, the door was left at four different increments of width for ten seconds and infiltration rates were measured. Figure 21 indicates a slower pace of infiltration when 25% of the entire doorway area was available for infiltration. The relationship between doorway area and the infiltration rate needs to be further investigated. During the first 10 seconds of 100% door opening period (100% of the door area was available for infiltration process), 22% of the cold air was infiltrated out of the cooler. This amount was reduced to 8% when 25% of the entire door area became available to the infiltration process. Looking at Figure 21, the curve that corresponds to the 100% door open shows a larger peak than when the door is opened/closed at 75% of the width. Both peaks are a direct result from the physical closing of the sliding door. The closing is most abrupt for the 100% of door width to keep the consistent 10-second time frame from one width to another. The flow of air in the cooler is governed by Navier-Stokes equations demonstrating elliptic behavior, therefore, any disturbance (like the door closing) in the flow is experienced through fluctuation in the tracer gas measurement. This weak disturbance is expected to be diminished and damped after a few seconds. Southern California Edison Page 41

56 FIGURE 21. INFILTRATION COMPARISON BETWEEN DIFFERENT DOOR POSITIONS (ADJACENT SPACE CONDITION OF 75 F DRY-BULB AND 55% RELATIVE HUMIDITY, AND FULL FAN SPEED) IMPACT OF COOLER S UTILIZATION VOLUME ON INFILTRATION As mentioned earlier, all models that were previously used did not consider the finite volume of the cooler and assumed a steady state solution. The tracer gas technique measured the concentration and the infiltration in real time, therefore captured the transient nature of the entire process. Subsequently, a parametric study/modeling was performed for one case to demonstrate the effect of the finite air volume inside the cooler. Figure 22 shows that if the cooler was filled with food products, less air would become available for infiltration causing a reduction in the infiltration rate. On the other hand, if the cooler was completely filled with food products there will not be any cold air available to be infiltrated out. Modeling results for this cooler showed that initial infiltration rates are approximately 4 ft 3 /sec for a 25% empty cooler, 17 ft 3 /sec for a 50% empty cooler, 38 ft 3 /sec for a 75% empty cooler, and 68 ft 3 /sec for a 100% empty cooler. Southern California Edison Page 42

57 *Infiltration rates are modeled from results using measured original cold air exfiltration methods FIGURE 22. INFILTRATION RATE AS A FUNCTION OF THE VOLUME OF COLD AIR INSIDE THE WALK-IN COOLER PARAMETRIC STUDIES (ADJACENT SPACE CONDITION OF 75 F DRY-BULB AND 55% RELATIVE HUMIDITY) VELOCITY PROFILE AND MAGNITUDE DURING INFILTRATION In this experiment, the PIV flow visualization technique was used to map the velocity profile and magnitude during infiltration. The air was seeded with smoke (very small particles) and the trajectory of the particles was captured through an interrogation window of 5 x 5 with a high speed camera. Two velocity components, horizontal (u) and vertical (v) were measured. The experiment was repeated to capture images along the height of the door opening in the mid-plane. All the images were assembled and the results are shown in Figure 23. Horizontal and vertical velocity contours are shown in Figure 24 and Figure 25, respectively. This image was consistent with the previous schematics shown in the ASHRAE (2006) handbook. The maximum velocity occurred at the lower section of the doorway where exfiltration from the walk-in cooler was occurring. The maximum velocity was about 65 centimeters/sec in the horizontal direction. It was interesting to see that the maximum vertical component occurred at the upper level of the doorway by the warm air that was entrained into the cooler. In Figure 23, affirming the conservation of mass, the outside warm air that infiltrated into the cooler from the upper portion of the door appears to be equal to the refrigerated air that exfiltrated into the ambient space from the lower portion of the door. Southern California Edison Page 43

58 The magnitude of the velocity components in Figure 24 and Figure 24 are shown with positive and negative values. For the horizontal portion (Figure 24), negative values coincide with infiltration, where positive values coincide with exfiltration. The vertical components (Figure 25) have positive values coincide with movement up and negative values coincide with movement down. FIGURE 23. VELOCITY VECTORS OR PROFILES DURING THE INFILTRATION Southern California Edison Page 44

59 FIGURE 24. HORIZONTAL VELOCITY CONTOURS DURING THE INFILTRATION (RIGHT +, LEFT -) FIGURE 25. VERTICAL VELOCITY CONTOURS DURING THE INFILTRATION (UP +, DOWN -) RESULTS OF VARYING EVAPORATOR FAN SPEED STUDY SINGLE REFRIGERATION COMPRESSOR SYSTEM FIXED CAPACITY Figure 26 illustrates the DB temperature and RH profiles in adjacent spaces during the entire test period for test scenario #14. Adjacent space conditions are illustrated with measurements of ambient conditions in rooms 2 and 4. Effectively, room 2 characterizes ambient conditions not directly exposed to exfiltration/door openings (see Figure 1: left adjacent space, near the closed hinged door), and room 4 characterizes ambient conditions that are directly exposed to exfiltration/door openings (see Figure 1: right adjacent space, near the sliding door). In either case, it is demonstrated that the adjacent space DB temperature was maintained at 80 F and RH at 60%, representative of kitchen-type adjacent space condition. It also serves as an example of how adjacent space conditions were maintained for all test scenarios over the entire test period. Temperature trend lines represent average temperatures across multiple sensors, indicating a representative room temperature at a given time during the test run. Relative humidities are reported from one sensor. Table 3 summarizes average DB temperatures and RH for all fixed capacity test scenarios conducted. As shown, target DB temperature and RH were maintained for all the test scenarios. Southern California Edison Page 45

60 4:15 PM 4:24 PM 4:33 PM 4:42 PM 4:51 PM 5:00 PM 5:09 PM 5:18 PM 5:27 PM 5:36 PM 5:45 PM 5:54 PM 6:03 PM 6:12 PM 6:21 PM 6:30 PM 6:39 PM 6:48 PM 6:57 PM 7:06 PM 7:15 PM 7:24 PM 7:33 PM 7:42 PM 7:51 PM 8:00 PM 8:09 PM Walk-in Surrounding Temperature ( F) [Room 2 and 4] Walk-in Surrounding Relative Humidity (%) [Room 2 and 4] Investigating the Effects of Fan Speed Variation and Rm 4 Avg Temp = 79.8 F Rm 2 Avg Temp = 80.1 F Rm 4 Avg RH = 60.1% Rm 2 Avg RH = 58.5% Defrost Period Entire Test Period (1-min data) (fixed capacity compressor % fan speed -- SCT 90 F) Average Temperature (Room 4, based on 11 sensors) Average Temperature (Room 2, based on 11 sensors) Relative Humidity (Room 4) Relative Humidity (Room 2) FIGURE 26. DRY-BULB TEMPERATURE AND RELATIVE HUMIDITY IN ADJACENT SPACES TABLE 3. SUMMARY OF MAINTAINED ADJACENT SPACE CONDITIONS FOR ALL FIXED CAPACITY TEST RUNS PARAMETER SCENARIO # FAN SPEED (%) TargetSCT ( F) Measured SCT ( F) Measured SET ( F) Rm 4 DB ( F) Rm 4 RH (%) Rm 2 DB ( F) Rm 2 RH (%) Figure 27 illustrates the trends in saturated evaporating temperature (SET) and saturated condensing temperature (SCT) for test scenario #14. Saturated temperatures were calculated values based on refrigerant properties. As a result, reported values during compressor off-cycles do not warrant any significant physical interpretations. The target SCT of 90 F was observed to be achieved in this scenario; Average compressor on-cycle SET was 16.8 F and average compressor on-cycle SCT was 90.1 F. Southern California Edison Page 46

61 FIGURE 27. KEY CONTROLLED VARIABLES A LOOK AT SATURATED EVAPORATING TEMPERATURES AND SATURATED CONDENSING TEMPERATURES Figure 28 shows all 12 product temperature profiles for the rack that was located on the opposite side of the fan coil. The product temperatures were observed to be within compliance of Food and Drug Administration s (FDA s) food code requirements of equal or below 41 F. In fact, these temperatures were observed to have the lowest product temperatures relative to other locations in the cooler. No significant changes in product temperature distribution were observed when cycling to 70% fan speed during compressor off cycles. In both scenarios, the left product on the 2 nd shelf was the warmest and the right product on the bottom shelf was the coldest product. Southern California Edison Page 47

62 FDA Food Code Requirement = 41 F Defrost Period FDA Food Code Requirement = 41 F Defrost Period FIGURE 28. OBSERVING THE IMPACTS OF VARYING FAN SPEED ON PRODUCT TEMPERATURES FOR A FIXED CAPACITY REFRIGERATION SYSTEM (PRODUCT RACK FACING THE FAN COIL EAST RACK) Figure 29 shows all 12 product temperature profiles for the rack that was located below the fan coil. In general, it was observed that product temperatures are higher in the rack below the fan coil than they are in the rack facing the fan coil. This is most likely due to the difficulty cold air would have in reaching the space directly Southern California Edison Page 48

63 below the coil versus directly in front of the coil. The product temperatures were observed to be within compliance of FDA food code requirements (41 F). No significant changes in product temperature distribution were observed when cycling to 70% fan speed during compressor off cycles. In both scenarios, the left product on the 3 rd shelf was the warmest and the center product on the bottom shelf was the coldest product. Product temperature changes are generally more pronounced in Figure 28 than in Figure 29. This is due to lower heat transfer coefficients associated with reduced airflow. Southern California Edison Page 49

64 FDA Food Code Requirement = 41 F Defrost Period FDA Food Code Requirement = 41 F FIGURE 29. OBSERVING THE IMPACTS OF VARYING FAN SPEED ON PRODUCT TEMPERATURES FOR A FIXED CAPACITY REFRIGERATION SYSTEM (PRODUCT RACK BELOW THE FAN COIL WEST RACK) Figure 30 shows all 12 product temperature profiles for the rack that was located along the sliding door. Only one of the product temperatures was observed to slightly exceed FDA food cod requirements in the base case scenario. The 70% fan speed scenario did not appear to be the cause of temperature slightly exceeding 41 F. No significant changes in product temperature distribution with 70% fan Southern California Edison Page 50

65 speed. In both scenarios, the right product on the 3 rd shelf was the warmest and the left product on the top shelf was the coldest product. FDA Food Code Requirement = 41 F FDA Food Code Requirement = 41 F Defrost Period FIGURE 30. OBSERVING THE IMPACTS OF VARYING FAN SPEED ON PRODUCT TEMPERATURES FOR A FIXED CAPACITY REFRIGERATION SYSTEM (RACK ALONG THE SLIDING DOOR NORTH RACK) Figure 31 shows all 12-product temperature profiles for the rack that was located along the swing or hinged type door. Product temperatures were observed to be Southern California Edison Page 51

66 within compliance of FDA food code requirements (41 F). No significant changes in product temperature distribution were observed when cycling to 70% fan speed during compressor off cycles. In both scenarios, the left product on the 3 rd shelf was the warmest and the right product on the 3 rd shelf was the coldest product. Also, the left product on the second shelf seemed to exhibit the same warm temperatures seen on the left product on the 3 rd shelf. Defrost Period Defrost Period Southern California Edison Page 52

67 FIGURE 31. OBSERVING THE IMPACTS OF VARYING FAN SPEED ON PRODUCT TEMPERATURES FOR A FIXED CAPACITY REFRIGERATION SYSTEM (RACK ALONG THE SWING TYPE DOOR SOUTH RACK) Some insight on the cold air supply/distribution throughout the walk-in cooler is apparent when comparing the trends observed in Figure 28, Figure 29, Figure 30, and Figure 31. It is observed that the most pronounced temperature pull-down is present in the product rack directly facing the evaporator. Conversely, the least pronounced product temperature pull-down is present in the product rack directly below the evaporator, and the product rack near the sliding door, i.e., cold air has the biggest challenge in reaching these areas: air does not have a direct path to the space right below the evaporator, and the cold air is supplied in a path that is closer to the rack near the sliding door than the rack near the swing-type door (in the fan coil, the fan closer to the swing door is running). Table 4 lists the average, minimum, and maximum product temperatures for all fixed capacity test runs. As shown, there were only a few instances that the maximum product temperature slightly exceeded 41 F, which was within experimental errors. This may also be attributed to the fact that some test scenarios reflect higher starting product temperatures than others, as these are not regulated to exact fixed levels prior to every test. TABLE 4. SUMMARY OF AVERAGE, MINIMUM, AND MAXIMUM PRODUCT TEMPERATURES FOR ALL FIXED CAPACITY TEST RUNS PARAMETER SCENARIO # FAN SPEED (%) Target SCT ( F) Avg. Product Temperatures ( F) Min. Product Temperatures ( F) Max. Product Temperatures ( F) Figure 32 depicts the temperature profiles along the height of the walk-in cooler. It also shows the RH measured at six-foot above the floor level. These measurements were taken near the swing type door or about 86 away from the sliding door. In either scenario, stratification did not seem to be an issue in the walk-in at ascending locations 86 away from the sliding door. Maximum temperature swings were approximately 6 F at 100% fan speed and 8 F at 70% fan speed. Similar observations were made about the temperature profiles near the sliding door. Figure 33 shows the temperature profiles along the height of the walk-in cooler and the RH measured at six-foot above the floor level. These measurements were taken at about 26 away from the sliding door. Stratification issues were not apparent in this location. Temp swings were approximately 16 F at this location. This, however, was expected because of the infiltration that takes place due to the door opening Southern California Edison Page 53

68 intervals. In other words, as the sliding door was opened, the warm air entered into the walk-in cooler and caused the air temperature to rise near the sliding door. Defrost Period Defrost Period FIGURE 32. OBSERVING THE IMPACTS OF VARYING FAN SPEED ON WALK-IN COOLER S INTERIOR TEMPERATURES FOR A FIXED CAPACITY REFRIGERATION SYSTEM (POLE 1 NEAR THE SWING TYPE DOOR) Southern California Edison Page 54

69 FIGURE 33. OBSERVING THE IMPACTS OF VARYING FAN SPEED ON WALK-IN COOLER S INTERIOR TEMPERATURES FOR A FIXED CAPACITY REFRIGERATION SYSTEM (POLE 2 NEAR THE SLIDING DOOR) Figure 34 shows the refrigerant mass flow rate and refrigeration effect for both test scenarios. As shown, in both scenarios, the refrigerant mass flow rate and refrigeration effect had same profiles. This demonstrated similar cooling loads for Southern California Edison Page 55

70 both scenarios. Although the refrigerant mass flow rate spiked higher in terms of instantaneous flow rates, it settled down roughly to the same average flow rate. Southern California Edison Page 56

71 FIGURE 34. OBSERVING THE REFRIGERATION EFFECT AND REFRIGERANT MASS FLOW RATE FOR A FIXED CAPACITY REFRIGERATION SYSTEM Figure 35 illustrates the power used by each component and the total power. The fan cycling trends are evident in Figure 35. As expected, the major contributor to the total power was the compressor. Since the fan power was two orders of magnitude small than the compressor power, it had a marginal impact on the total power. As a result, contribution of varying fan power to the total cooling load in the walk-in was indistinguishable. Southern California Edison Page 57

72 FIGURE 35. ANALYZING THE EFFECTS OF VARYING FAN SPEED ON POWER CONSUMPTION FOR A FIXED CAPACITY REFRIGERATION SYSTEM Figure 36 shows the energy by each component and total energy during the refrigeration period over the test cycle. There was no significant difference between the compressor energy with and without VSD fan. The difference was only 0.04 kwh or less than 2% per test cycle, which was well within the experimental errors. The decrease in fan energy, on the other hand, was about 0.03 kwh or 30% per test Southern California Edison Page 58

73 cycle. As expected, there was no change in lighting energy usage. No substantial impact, however, was observed on total energy usage. FIGURE 36. ANALYZING THE EFFECTS OF VARYING FAN SPEED ON ENERGY CONSUMPTION FOR A FIXED CAPACITY REFRIGERATION SYSTEM MULTIPLEX REFRIGERATION COMPRESSOR SYSTEM VARIABLE CAPACITY The results for multiplex refrigeration compressor test runs are divided into three scenarios or categories. These three scenarios are: Scenario A cooling load was varied while keeping fan at fixed full speed Scenario B fan speed was varied with variations in cooling load Scenario C fan speed was varied while keeping a fixed high cooling load SCENARIO A VARYING THE COOLING LOAD WHILE KEEPING FAN AT FIXED FULL SPEED Figure 37 through Figure 40 depict the product temperature profiles for all four racks in the walk-in cooler. All 48 product temperatures were observed to be within the FDA food code requirements (41 F). Although no significant changes in product temperature distributions were observed, product temperature under low load conditions seemed to be lower in temperatures than those under high load conditions. Southern California Edison Page 59

74 FIGURE 37. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK FACING THE FAN COIL EAST RACK, SCENARIO A) Southern California Edison Page 60

75 FDA Food Code Requirement = 41 F FIGURE 38. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK BELOW THE FAN COIL WEST RACK, SCENARIO A) Southern California Edison Page 61

76 FDA Food Code Requirement = 41 F FIGURE 39. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK ALONG THE SLIDING DOOR NORTH RACK, SCENARIO A) Southern California Edison Page 62

77 FIGURE 40. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK ALONG THE SWING TYPE DOOR SOUTH RACK, SCENARIO A) Figure 41 depicts the temperature profiles along the height of the walk-in cooler. It also shows the RH measured at six-foot above the floor level. These measurements were taken near the swing type door or about 86 away from the sliding door. In either scenario, stratification did not seem to be an issue in the walk-in at ascending Southern California Edison Page 63

78 locations 86 away from the sliding door. Due to the door openings, temperature variations were observed under high load condition. Defrost Period FIGURE 41. OBSERVING THE WALK-IN COOLER S INTERIOR TEMPERATURES FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (POLE 1 NEAR THE SWING TYPE DOOR, SCENARIO A) Similar observations were made about the temperature profiles near the sliding door. Figure 42 shows the temperature profiles along the height of the walk-in cooler and Southern California Edison Page 64

79 the RH measured at six-foot above the floor level. These measurements were taken at about 26 away from the sliding door. Stratification issues were not apparent in this location. Due to the door openings, however, temperature variations were observed under high load condition. Southern California Edison Page 65

80 FIGURE 42. OBSERVING THE WALK-IN COOLER S INTERIOR TEMPERATURES FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (POLE 2 NEAR THE SLIDING DOOR, SCENARIO A) Figure 43 illustrates the refrigerant mass flow rate and refrigeration effect. As illustrated, both parameters had similar profiles. The refrigerant mass flow rate profile under low load conditions, however, seemed to have more regular patterns. FIGURE 43. OBSERVING THE REFRIGERATION EFFECT AND REFRIGERANT MASS FLOW RATE FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO A) Southern California Edison Page 66

81 Figure 44 shows the power used by each component and the total power. As shown, the major contributor to the total power was the compressor. Also, the compressor power profile under low load condition demonstrates the compressor cycled more under this condition compared to high load condition. FIGURE 44. OBSERVING THE POWER CONSUMPTION TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO A) Southern California Edison Page 67

82 Figure 45 shows the energy by each component and total energy during the refrigeration period over the test cycle. As shown, the compressor energy usage was dropped by more than 10% under low load condition. However, the lighting and fan energy usage remained unchanged. FIGURE 45. OBSERVING THE ENERGY CONSUMPTION FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO A) SCENARIO B VARYING THE FAN SPEED WITH COOLING LOAD VARIATION Figure 46 through Figure 49 illustrate the product temperature profiles for all four racks in the walk-in cooler. All 48 product temperatures were observed to be within the FDA food code requirements (41 F). Although no significant changes in product temperature distributions were observed, product temperature under low load and low speed fan conditions seemed to be lower in temperatures than those under high load and full speed fan condition. Southern California Edison Page 68

83 FDA Food Code Requirement = 41 F FIGURE 46. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK FACING THE FAN COIL EAST RACK, SCENARIO B) Southern California Edison Page 69

84 FDA Food Code Requirement = 41 F FIGURE 47. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK BELOW THE FAN COIL WEST RACK, SCENARIO B) Southern California Edison Page 70

85 FDA Food Code Requirement = 41 F FIGURE 48. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK ALONG THE SLIDING DOOR NORTH RACK, SCENARIO B) Southern California Edison Page 71

86 FDA Food Code Requirement = 41 F FIGURE 49. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK ALONG THE SWING TYPE DOOR SOUTH RACK, SCENARIO B) Southern California Edison Page 72

87 Figure 50 shows the temperature profiles along the height of the walk-in cooler. It also shows the RH measured at six-foot above the floor level. These measurements were taken near the swing type door or about 86 away from the sliding door. In either scenario, stratification did not seem to be an issue in the walk-in at ascending locations 86 away from the sliding door. Due to the door openings, temperature variations were observed under high load and full speed fan condition. Defrost Period FIGURE 50. OBSERVING THE WALK-IN COOLER S INTERIOR TEMPERATURES FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (POLE 1 NEAR THE SWING TYPE DOOR, SCENARIO B) Southern California Edison Page 73

88 Figure 51 shows the temperature profiles along the height of the walk-in cooler and the RH measured at six-foot above the floor level. These measurements were taken at about 26 away from the sliding door. As shown, stratification issues were not apparent in this location. Due to the door openings, however, temperature variations were observed under high load and full speed fan condition. Southern California Edison Page 74

89 FIGURE 51. OBSERVING THE WALK-IN COOLER S INTERIOR TEMPERATURES FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (POLE 2 NEAR THE SLIDING DOOR, SCENARIO B) Figure 52 illustrates the refrigerant mass flow rate and refrigeration effect. Similar profiles were observed for both parameters under both test conditions. Although no significant differences were observed, more regular refrigerant flow patterns were observed low load and low fan speed. Southern California Edison Page 75

90 FIGURE 52. OBSERVING THE REFRIGERATION EFFECT AND REFRIGERANT MASS FLOW RATE FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO B) Figure 53 shows the power used by each component and the total power. As shown, the major contributor to the total power was the compressor. Also, the compressor power profile under low load and low fan speed condition demonstrated that the compressor cycled more under this condition compared to high load and full speed fan condition. Southern California Edison Page 76

91 FIGURE 53. OBSERVING THE POWER CONSUMPTION TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO B) Figure 54 shows the energy by each component and total energy during the refrigeration period over the test cycle. As shown, the compressor energy usage was dropped by about 13% under low load and low fan speed condition. Also, the fan Southern California Edison Page 77

92 energy was dropped by about 50% under low load and low fan speed condition. As expected, the lighting energy usage did not change. FIGURE 54. OBSERVING THE ENERGY CONSUMPTION FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO B) SCENARIO C VARYING THE FAN SPEED WHILE KEEPING A FIXED HIGH COOLING LOAD Figure 55 through Figure 58 depicts the product temperature profiles for all four racks in the walk-in cooler. With the exception of one location under low fan speed condition, the remaining product temperatures were observed to be within the FDA food code requirements (41 F). The only location that the temperature was observed to be above 41 F was on the rack along the sliding door, top shelf right location (Figure 57). It is also worth to note that even with the fan funning at full speed, the product temperatures were close to 41 F. Southern California Edison Page 78

93 FDA Food Code Requirement = 41 F FIGURE 55. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK FACING THE FAN COIL EAST RACK, SCENARIO C) Southern California Edison Page 79

94 FDA Food Code Requirement = 41 F FIGURE 56. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK BELOW THE FAN COIL WEST RACK, SCENARIO C) Southern California Edison Page 80

95 FDA Food Code Requirement = 41 F FIGURE 57. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK ALONG THE SLIDING DOOR NORTH RACK, SCENARIO C) Southern California Edison Page 81

96 FDA Food Code Requirement = 41 F FIGURE 58. OBSERVING THE PRODUCT TEMPERATURE TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (RACK ALONG THE SWING TYPE DOOR SOUTH RACK, SCENARIO C) Figure 59 and Figure 60 show the temperature profiles along the height of the walkin cooler at 86 and 26 away from the sliding door, respectively. These figures also show the RH inside the walk-in measured at six-foot above the floor level. As shown, although the temperatures varied during the door opening intervals, there were no Southern California Edison Page 82

97 stratification along the height of the walk-in cooler. Clearly, the temperature measurements that were closer to the sliding door were more vulnerable to adjacent space conditions, and therefore experienced higher temperature swings. Defrost Period FIGURE 59. OBSERVING THE WALK-IN COOLER S INTERIOR TEMPERATURES FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (POLE 1 NEAR THE SWING TYPE DOOR, SCENARIO C) Southern California Edison Page 83

98 FIGURE 60. OBSERVING THE WALK-IN COOLER S INTERIOR TEMPERATURES FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (POLE 2 NEAR THE SLIDING DOOR, SCENARIO C) Figure 61 illustrates the refrigerant mass flow rate and refrigeration effect. Similar profiles were observed for both parameters under both test conditions. Although no Southern California Edison Page 84

99 significant differences were observed, more regular refrigerant flow patterns were observed low fan speed condition. FIGURE 61. OBSERVING THE REFRIGERATION EFFECT AND REFRIGERANT MASS FLOW RATE FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO C) Figure 62 shows the power used by each component and the total power. As shown, the major contributor to the total power was the compressor. Under low fan speed condition, less fluctuations in compressor power was observed. Southern California Edison Page 85

100 FIGURE 62. OBSERVING THE POWER CONSUMPTION TRENDS FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO C) Figure 63 shows the energy by each component and total energy during the refrigeration period over the test cycle. There was no significant difference between the compressor energy with and without VSD fan. The difference was only 0.04 kwh or less than 1% per test cycle, which was well within the experimental errors. The Southern California Edison Page 86

101 decrease in fan energy, on the other hand, was about 0.06 kwh or 50% per test cycle. As expected, there was no change in lighting energy usage. FIGURE 63. OBSERVING THE ENERGY CONSUMPTION FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM (SCENARIO C) Table 5 and Table 6 summarized key parameters for all 12 test runs for variable capacity refrigeration systems. TABLE 5. AVERAGE SATURATED EVAPORATING/CONDENSING TEMPERATURES, AVERAGE PRODUCT TEMPERATURES, AND AVERAGE ADJACENT SPACE CONDITIONS FOR VARIABLE CAPACITY TEST RUNS Parameter Scenario # Evaporator Load High High High High High High FAN SPEED (%) Target SCT ( F) Measured SCT ( F) Southern California Edison Page 87

102 Parameter Measured SET ( F) Avg. Product Temperatures ( F) Min. Product Temperatures ( F) Max. Product Temperatures ( F) Scenario # Rm 4 DB ( F) Rm 4 RH (%) Rm 2 DB ( F) Rm 2 RH (%) TABLE 6. AVERAGE SATURATED EVAPORATING/CONDENSING TEMPERATURES, AVERAGE PRODUCT TEMPERATURES, AND AVERAGE ADJACENT SPACE CONDITIONS FOR VARIABLE CAPACITY TEST RUNS (CONTINUED) Parameter Evaporator Load FAN SPEED (%) Target SCT ( F) Measured SCT ( F) Measured SET ( F) Avg. Product Temperatures ( F) Scenario # High High High High High High Southern California Edison Page 88

103 Parameter Min. Product Temperatures ( F) Max. Product Temperatures ( F) Scenario # Rm 4 DB ( F) Rm 4 RH (%) Rm 2 DB ( F) Rm 2 RH (%) IMPACT OF FLOATING HEAD PRESSURES FIXED CAPACITY REFRIGERATION SYSTEM Figure 64 shows that SCT were maintained at fixed levels for full speed fan test scenario. This is an example of how SCT s were maintained. Southern California Edison Page 89

104 FIGURE 64. IMPACTS OF FLOATING HEAD PRESSURE ON SATURATED EVAPORATING TEMPERATURE AND SATURATED CONDENSING TEMPERATURE FOR A FIXED CAPACITY REFRIGERATION SYSTEM Figure 65 and Figure 66 illustrate the trends in power consumption by each component for full speed and 70% fan speed test runs, respectively. As illustrated, the power demand increased as SCT was increased. Also, the total power demand did not change significantly according to changes in fan speed. Southern California Edison Page 90

105 FIGURE 65. IMPACTS OF FLOATING HEAD PRESSURE ON POWER CONSUMPTION FOR A FIXED CAPACITY REFRIGERATION SYSTEM Southern California Edison Page 91

106 FIGURE 66. IMPACTS OF FLOATING HEAD PRESSURE AND VARIABLE SPEED EVAPORATOR FAN CONTROLS ON POWER CONSUMPTION FOR A FIXED CAPACITY REFRIGERATION SYSTEM Figure 66 and Figure 68 depict the energy usage by each component for refrigeration test period for full speed and 70% fan speed test runs, respectively. As illustrated, the energy usage increased as SCT was increased. Also, the total energy usage did not change significantly according to changes in fan speed. In addition, the main energy usage were due to the compressor. Southern California Edison Page 92

107 FIGURE 67. IMPACTS OF FLOATING HEAD PRESSURE ON ENERGY CONSUMPTION FOR A FIXED CAPACITY REFRIGERATION SYSTEM FIGURE 68. IMPACTS OF FLOATING HEAD PRESSURE AND VARIABLE SPEED EVAPORATOR FAN CONTROLS ON ENERGY CONSUMPTION FOR A FIXED CAPACITY REFRIGERATION SYSTEM Southern California Edison Page 93

108 VARIABLE CAPACITY REFRIGERATION SYSTEM Figure 69 shows that SCT were maintained at fixed levels for full speed fan and high load test scenarios. This is an example of how SCT s were maintained. FIGURE 69. IMPACTS OF FLOATING HEAD PRESSURE ON SATURATED EVAPORATOR TEMPERATURE AND SATURATED CONDENSING TEMPERATURE FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM Figure 70 and Figure 71 illustrate the trends in power consumption by each component for full speed and 70% fan speed test runs, respectively. As illustrated, Southern California Edison Page 94

109 the power demand increased as SCT was increased. Also, the total power demand did not change significantly according to changes in fan speed. FIGURE 70. IMPACTS OF FLOATING HEAD PRESSURE ON POWER CONSUMPTION FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM Southern California Edison Page 95

110 FIGURE 71. IMPACTS OF FLOATING HEAD PRESSURE AND A VARIABLE SPEED EVAPORATOR FAN ON POWER CONSUMPTION FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM Figure 72 and Figure 73 depict the energy usage by each component for refrigeration test period for full speed and 70% fan speed test runs, respectively. As illustrated, the energy usage increased as SCT was increased for both cases. The total energy usage under high load condition, however, was higher than that for the low load condition. In addition, the main energy usage were due to the compressor. Southern California Edison Page 96

111 FIGURE 72. IMPACTS OF FLOATING HEAD PRESSURE ON ENERGY CONSUMPTION FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM FIGURE 73. IMPACTS OF FLOATING HEAD PRESSURE AND A VARIABLE SPEED EVAPORATOR FAN ON ENERGY CONSUMPTION FOR A VARIABLE CAPACITY REFRIGERATION SYSTEM Southern California Edison Page 97

112 CONCLUSIONS The following lists the conclusions drawn from this project: The infiltration rate is a time-dependent process that follows a decaying function behavior as conditions between air inside the walk-in and air in adjacent space reach equilibrium. Before coming to equilibrium, 75% of infiltration occurs (75% of cold air in the cooler is lost) during the first 30 seconds of a door opening. Equilibrium is reached (all cold air in the cooler is lost) at 120 seconds of a door opening. The infiltration rate is also dependent on how much product the walk-in cooler contains. An empty cooler has higher volumes of cold air that may be lost. Modeling results for this cooler showed that initial infiltration rates are approximately 4 ft 3 /sec for a 25% empty cooler, 17 ft 3 /sec for a 50% empty cooler, 38 ft 3 /sec for a 75% empty cooler, and 68 ft 3 /sec for a 100% empty cooler. The current industry-wide methodologies for estimating the infiltration rate in their steady state form are not sufficient for predictions. They over-estimate the infiltration rate for durations of more than seconds. Even after modifying for time-varying densities, they still over-predicted infiltration and the cooling load thereafter. The fan speed does not impact the infiltration rate for small temperature differences between the refrigerated space and adjacent warm space. During the period of door opening, for a temperature difference of about 45 F turning the fan off from running at full speed will lower the infiltration by about 5-7%. However, as this temperature difference increases to 80 F, the reduction in infiltration will be about 15-20% during the first 20 seconds of the infiltration process. The flow visualization technique of PIV mapped the velocity profile and proved that the schematics currently used to demonstrate air movement at the door opening capture the direction of motion. However, the values of the discharge velocity components have never been measured by the state-of-the-art PIV technology. There were no significant changes or impacts on the product temperature maintenance as well as room temperature stratification as a function of fan speed. The cooling load was the most dominant factor in power demand and energy usage compared to evaporator fan speed variations. This indicated that the heat gain from the ECM fan motor had little to no impact on the total cooling load due to its low power consumption. Floating heat pressure had more substantial impact on power demand and energy consumption than the variable evaporator fan speed. This was observed in both fixed and variable capacity refrigeration systems. Southern California Edison Page 98

113 RECOMMENDATIONS To understand the full impact of individual parameters on infiltration, they must be adjusted singularly. Different fan placement is recommended as a new parameter that needs to be tested with CFD and a modular scaled model. Although the model developed in this project links the volume of air inside the cooler to the infiltrated amount of air, more tests with several filled volumes of the cooler are recommended for further verification. To confirm that there are no significant differences between the door type (swing vs. sliding) on the infiltration rate, testing at the sliding door needs to be duplicated for swing or hinged type doors. Using PIV at the midway plane of the swing type door would further verify this claim. In addition, to confirm and duplicate the impacts of variable speed evaporator fan motors in walk-ins, it is recommended to further investigate the fan speed variations in walk-in freezer applications as well as larger walk-in sizes and configurations. Southern California Edison Page 99

114 APPENDIX A CALCULATIONS OF MIXTURE DENSITY OF AIR The mixture of gases is composed of CO2, dry air (da), and water vapor (v). The mixture density is given as: mix i i i where C mix i i C 3 Total effective density, accounting for each constituent, (kg/m ) 3 Density of each constituent: CO 2, da, v,(kg/m ) Mass fraction of each constituent: CO, da, v The density of each constituent is obtained from the ideal gas law with the corresponding gas constants as: CO P T P v 0.462T P 0.287T da Where P = Pressure (kpa) T = Temperature, ( K) 2 (A-1) (A-2) The mass fraction of each constituent is directly correlated to the mole fraction. If y represents the mole fraction, ( ) as a function of time t is known. The mole fraction of y CO t 2 CO2 in time can be obtained from Figures 5, 7, 8, and 9 (or similar figures). The humidity ratio can be obtained from the temperature and RH. Let us designate the symbol to humidity ratio. So: C C v da 1 (A-3) 1 C v The mole fraction of water vapor and dry air can be obtained as: Southern California Edison Page 100

115 n n n y y v da total ' v ' da Cv 18 Cda n n n n v v total n n da total da (A-4) However, these mole fractions are based on 1 mole of the dry air and vapor mixture. We have only 1 ( t) moles available, therefore the mole fraction of each constituent 2 becomes: y y CO 2 ( t) y ( t) v da ( t) y ' v y 1 ' da 1 y CO y CO y 2 ( t) CO 2 ( t) (A-5) The mass fractions of each constituent can be obtained according to the following formulas: m C total CO2 ( t) y 44 CO2 y ( t) CO2 ( t) m total y ( t) v 18 y da ( t) yv ( t) 18 (A-6) Cv ( t) mtotal yda ( t) Cda ( t) mtotal Note that: M Molecular weight in kg / kmol K M M M CO2 v da The specific heat of the mixture to be used in Gosney s equation (2) can be calculated as: cp mix i c pi y i (A-7) The enthalpy difference for Equation 2 is calculated as: h i h r c p mix T i T r Where i and r refer to the inside (inside walk in cooler) and outside (adjacent space) conditions, respectively. (A-8) Southern California Edison Page 101

116 APPENDIX B NATURAL INFILTRATION THROUGH UNSEALED AREAS The test conditions and the frequency of data collected for natural infiltration (or decay of CO 2 ) concentration inside the cooler is shown in Table B-1. TABLE B-1 THE ORIGINAL NUMBER OF DATA POINTS FROM CAPTURED NATURAL DECAY OF THE TEMPERATURE/RELATIVE HUMIDITY PROFILES AT DIFFERENT FAN SPEEDS Type of Test - Fan Speed (DB/RH) Original Number of Data Points (1 data point per second) decay - 100% fan (75/55) 2320 decay - 75% fan (75/55) 800 decay - 50% fan (75/55) 660 decay - 100% fan (80/60) 1998 decay - 75% fan (80/60) 458 decay - 50% fan (80/60) 458 decay - 100% fan (84/82) 960 decay - 75% fan (84/82) 914 decay - 50% fan (84/82) 540 decay - 100% fan (115/14) 1250 decay - 75% fan (115/14) 600 Southern California Edison Page 102

117 APPENDIX C FINDING TRACER GAS CONCENTRATION GRADIENT IN TIME The logged data shown in Figures C-1 through C-4, depicts the exfiltration of the walk-in cooler air tagged with tracer gas. The graphs show the normalized concentration during and after the period the door is left open (3, 30, 120, and 600 seconds). The concentration is continuously monitored to steady state by mixing after the door is closed. The difference in concentration represents the infiltration during the related time period, i.e., 3, 30, 120, and 600 seconds. The concentration change for each curve corresponds to the duration of actual infiltration period (3, 30, 120, etc. seconds). Curves in each group are for different fan speeds. Subsequently, a file was created illustrating time on the horizontal axis (x-axis) and the concentration change on the vertical axis (y-axis), read from Figures C-1 through C-4. These files create the tracer gas concentration gradient in time shown in Equation 4. Note that the volume fraction represents the mole fraction. Southern California Edison Page 103

118 FIGURE C-1 THE CONCENTRATION VARIATION BEFORE AND AFTER THE INFILTRATION FOR SEVERAL PERIODS OF TIME THAT THE DOOR STANDS OPEN-75 F AND 55% RH FIGURE C-2 THE CONCENTRATION VARIATION BEFORE AND AFTER THE INFILTRATION FOR SEVERAL PERIODS OF TIME THAT THE DOOR STANDS OPEN - 80 F AND 60% RH Southern California Edison Page 104

119 FIGURE C-3 THE CONCENTRATION VARIATION BEFORE AND AFTER THE INFILTRATION FOR SEVERAL PERIODS OF TIME THAT THE DOOR STANDS OPEN - 84 F AND 82% RH FIGURE C-4 THE CONCENTRATION VARIATION BEFORE AND AFTER THE INFILTRATION FOR SEVERAL PERIODS OF TIME THAT THE DOOR STANDS OPEN F AND 14% RH Southern California Edison Page 105

120 APPENDIX D CFD ANALYSIS In order to identify the regions in the walk-in cooler (without any food inside) when one fan is running and the doors are closed, the temperatures of different points inside the cooler were tracked with time. An initial uniform temperature of 25 o C was cooled down after the fan started blowing cool air coming from the evaporator. Figure D-1 presents the 14 points that the temperatures are tracked with time. FIGURE D-1, LOCATION OF TRACKING POINTS In the CFD simulation the Reynolds Stress turbulent model was utilized. The contour plots of temperatures at some selected sections are shown in Figure D-2 and D-3. The selected section of Figure D-2 passes vertically through the center of the running fan, while in Figure D-3, one plane passes vertically through the running fan and the other one is horizontally positioned 6.6 ft. (2 m) below the fans. Southern California Edison Page 106

121 FIGURE D-2 TEMPERATURE CONTOURS AT THE EVAPORATOR EXIT AND AT THE PLANE PASSING VERTICALLY THROUGH THE RUNNING FAN. FIGURE D-3 TEMPERATURE CONTOURS AT THE EXIT OF THE EVAPORATOR AND AT THE PLANE PASSING HORIZONTALLY THROUGH THE RUNNING FAN AND 6.6 FT (2 M) BELOW. The results indicate that the corner located near the ceiling and on the fan side has the lowest temperature, and the point near the floor on the opposite side of the running fan experiences the highest temperature. Southern California Edison Page 107

122 APPENDIX E INFILTRATION RATES FOR VARIOUS ADJACENT SPACE CONDITIONS Adjacent Space Condition: 75 o F and 55%RH FIGURE E1. TAMM S ORIGINAL EQUATION [CONSTANT DENSITY] (LEFT) AND MODIFIED EQUATION [VARIABLE DENSITY] (RIGHT) Southern California Edison Page 108

123 FIGURE E2. GOSNEY S ORIGINAL EQUATION [CONSTANT DENSITY] (LEFT) AND MODIFIED EQUATION [VARIABLE DENSITY] (RIGHT) FIGURE E3. CLELAND S ORIGINAL EQUATION WITH CONSTANT DENSITY (LEFT) AND WITH VARIABLE DENSITY (RIGHT) FIGURE E4. CLELAND S MODIFIED EQUATION WITH CONSTANT DENSITY [ADJUSTED FOR ABRUPT EXPANSION] Southern California Edison Page 109

124 FIGURE E5. DENSITY VARIATION INSIDE AND OUTSIDE THE COOLER AT DIFFERENT FAN SPEEDS Adjacent Space Condition: 80 o F and 60%RH Southern California Edison Page 110

125 FIGURE E6. TAMM S ORIGINAL EQUATION [CONSTANT DENSITY] (LEFT) AND MODIFIED EQUATION [VARIABLE DENSITY] (RIGHT) FIGURE E7. GOSNEY S ORIGINAL EQUATION [CONSTANT DENSITY] (LEFT) AND MODIFIED EQUATION [VARIABLE DENSITY] (RIGHT) FIGURE E8. CLELAND S ORIGINAL EQUATION WITH CONSTANT DENSITY (LEFT) AND WITH VARIABLE DENSITY (RIGHT) Southern California Edison Page 111

126 FIGURE E9. CLELAND S MODIFIED EQUATION WITH CONSTANT DENSITY [ADJUSTED FOR ABRUPT EXPANSION] FIGURE E10. DENSITY VARIATION INSIDE AND OUTSIDE THE COOLER AT DIFFERENT FAN SPEEDS Adjacent Space Condition: 84 o F and 82%RH Southern California Edison Page 112

127 FIGURE E11. TAMM S ORIGINAL EQUATION [CONSTANT DENSITY] (LEFT) AND MODIFIED EQUATION [VARIABLE DENSITY] (RIGHT) Southern California Edison Page 113

128 FIGURE E12. GOSNEY S ORIGINAL EQUATION [CONSTANT DENSITY] (LEFT) AND MODIFIED EQUATION [VARIABLE DENSITY] (RIGHT) FIGURE E13. CLELAND S ORIGINAL EQUATION WITH CONSTANT DENSITY (LEFT) AND WITH VARIABLE DENSITY (RIGHT) FIGURE E14. CLELAND S MODIFIED EQUATION WITH CONSTANT DENSITY [ADJUSTED FOR ABRUPT EXPANSION] Southern California Edison Page 114

129 FIGURE E15. DENSITY VARIATION INSIDE AND OUTSIDE THE COOLER AT DIFFERENT FAN SPEEDS Adjacent Space Condition: 115 o F and 14%RH FIGURE E16. TAMM S ORIGINAL EQUATION [CONSTANT DENSITY] (LEFT) AND MODIFIED EQUATION [VARIABLE DENSITY] (RIGHT) Southern California Edison Page 115

130 FIGURE E17. GOSNEY S ORIGINAL EQUATION [CONSTANT DENSITY] (LEFT) AND MODIFIED EQUATION [VARIABLE DENSITY] (RIGHT) FIGURE E18. CLELAND S ORIGINAL EQUATION WITH CONSTANT DENSITY (LEFT) AND WITH VARIABLE DENSITY (RIGHT) Southern California Edison Page 116

131 FIGURE E19. CLELAND S MODIFIED EQUATION WITH CONSTANT DENSITY [ADJUSTED FOR ABRUPT EXPANSION] FIGURE E20. DENSITY VARIATION INSIDE AND OUTSIDE THE COOLER AT DIFFERENT FAN SPEEDS Infiltration rate is the amount of warm air going into the cooler and is equal to the exfiltration rate - that is the amount of cold air leaving the cooler (concept of infiltration/exfiltration is based on conservation of mass) therefore, we use these terms interchangeably. The infiltration rate is the amount of warm air going into the cooler in time. This rate is a function of time as the amount of cold air to be exfiltrated to the outside room reduces in Southern California Edison Page 117

132 time. Using Figure E21 as a typical measurement of the infiltration rate, we can state that if a unit volume of infiltrated air is considered at two different time intervals, they do not contain the same amount of cold air. The pace by which cold air that is moving outside the cooler changes with time. For instance at 60 seconds, the infiltration rate of cold air moving out or cold air is 3 ft³/sec. *Infiltration rates are based on methods using measured original cold air exfiltration FIGURE E21 A TYPICAL INFILTRATION RATE CURVE OBTAINED BY MEASUREMENTS By referring to Figure E21, it can be seen that the infiltration rate goes to 0 as time (t) approaches. The infiltration rate asymptotically goes to zero. If the door is left opened for a long time (say for example 10 minutes), the entire cold air leaves the cooler and is therefore replace by warm air (as the rate of discharge goes to zero). This is different from the total amount of air (mass or volume) that is displaced. This amount will be the area under the infiltration rate curve. Figure C-1 represents the total amount of cold air left in the room. With reference to Figure E21, let s pick an arbitrary time (t 1 ) during which the door is left opened. We have representing the volume and representing volume per unit time (volumetric infiltration/exfiltration rate): Southern California Edison Page 118

INTEGRATION OF DEMAND RESPONSE INTO TITLE 20 FOR COMMERCIAL ICE MACHINES

INTEGRATION OF DEMAND RESPONSE INTO TITLE 20 FOR COMMERCIAL ICE MACHINES Design & Engineering Services INTEGRATION OF DEMAND RESPONSE INTO TITLE 20 FOR COMMERCIAL ICE MACHINES Phase1: Demand Response Potential DR 09.05.06 Report Prepared by: Design & Engineering Services Customer

More information

Hot Food Holding Cabinets Test for FWE HLC

Hot Food Holding Cabinets Test for FWE HLC Design & Engineering Services Hot Food Holding Cabinets Test for Report Prepared by: Design & Engineering Services Customer Service Business Unit Southern California Edison January 4, 2011 Acknowledgements

More information

PPL Electric Utilities Energy Efficiency and Conservation Program

PPL Electric Utilities Energy Efficiency and Conservation Program Installation of High-Efficiency High Efficiency Evaporator Fans Evaporator Fans That have Highfor Walk-ins Efficiency Permanent Split $50/fan motor Capacitor (PSC) Motor or ECM High Efficiency Compressors

More information

INTEGRATION OF DEMAND RESPONSE INTO TITLE 20 FOR HOT FOOD HOLDING CABINETS

INTEGRATION OF DEMAND RESPONSE INTO TITLE 20 FOR HOT FOOD HOLDING CABINETS Design & Engineering Services INTEGRATION OF DEMAND RESPONSE INTO TITLE 20 FOR HOT FOOD HOLDING CABINETS Phase1: Demand Response Potential DR 09.05.07 Report Prepared by: Design & Engineering Services

More information

CONVECTION OVENS FOR FOOD SERVICE APPLICATIONS BLODGETT SHO-E

CONVECTION OVENS FOR FOOD SERVICE APPLICATIONS BLODGETT SHO-E Design & Engineering Services CONVECTION OVENS FOR FOOD SERVICE APPLICATIONS BLODGETT SHO-E Report ` Prepared by: Design & Engineering Services Customer Service Business Unit Southern California Edison

More information

Open and Closed Door Moisture Transport and Corresponding Energy Consumption in Household Refrigerator

Open and Closed Door Moisture Transport and Corresponding Energy Consumption in Household Refrigerator 18 R. Saidur et al./journal of Energy & Environment, Vol. 6, May 2007 Open and Closed Door Moisture Transport and Corresponding Energy Consumption in Household Refrigerator R. Saidur, M. A. Sattar, M.

More information

WHITE PAPER. ANSI/AHRI Standard for Fan and Coil Evaporators - Benefits and Costs

WHITE PAPER. ANSI/AHRI Standard for Fan and Coil Evaporators - Benefits and Costs Abstract Fan and coil evaporators as used in the industrial refrigeration industry can be certified for performance per ANSI/AHRI Standard 420-2008, Performance Rating of Forced-Circulation Free-Delivery

More information

PRESSURE-ENTHALPY CHARTS AND THEIR USE By: Dr. Ralph C. Downing E.I. du Pont de Nemours & Co., Inc. Freon Products Division

PRESSURE-ENTHALPY CHARTS AND THEIR USE By: Dr. Ralph C. Downing E.I. du Pont de Nemours & Co., Inc. Freon Products Division INTRODUCTION PRESSURE-ENTHALPY CHARTS AND THEIR USE The refrigerant in a refrigeration system, regardless of type, is present in two different states. It is present as liquid and as vapor (or gas). During

More information

VARIABLE REFRIGERANT FLOW FOR LODGING

VARIABLE REFRIGERANT FLOW FOR LODGING Design & Engineering Services VARIABLE REFRIGERANT FLOW FOR LODGING APPLICATIONS Report Prepared by: Design & Engineering Services Customer Service Business Unit Southern California Edison June 2012 Acknowledgements

More information

ME 410 MECHANICAL ENGINEERING SYSTEMS LABORATORY MASS & ENERGY BALANCES IN PSYCHROMETRIC PROCESSES EXPERIMENT 3

ME 410 MECHANICAL ENGINEERING SYSTEMS LABORATORY MASS & ENERGY BALANCES IN PSYCHROMETRIC PROCESSES EXPERIMENT 3 ME 410 MECHANICAL ENGINEERING SYSTEMS LABORATORY MASS & ENERGY BALANCES IN PSYCHROMETRIC PROCESSES EXPERIMENT 3 1. OBJECTIVE The objective of this experiment is to observe four basic psychrometric processes

More information

FS 231: Final Exam (5-6-05) Part A (Closed Book): 60 points

FS 231: Final Exam (5-6-05) Part A (Closed Book): 60 points Name: Start time: End time: FS 231: Final Exam (5-6-05) Part A (Closed Book): 60 points 1. What are the units of the following quantities? (10 points) a. Enthalpy of a refrigerant b. Dryness fraction of

More information

Energy Transfer Based Test Method Development and Evaluation of Horizontal Air Flow Re- Circulatory Air Curtain Efficiencies

Energy Transfer Based Test Method Development and Evaluation of Horizontal Air Flow Re- Circulatory Air Curtain Efficiencies Purdue University Purdue e-pubs International High Performance Buildings Conference School of Mechanical Engineering 2012 Energy Transfer Based Test Method Development and Evaluation of Horizontal Air

More information

A/C Cooling Load calculation and measurement

A/C Cooling Load calculation and measurement Testo Inc. 40 White Lake Rd. Sparta NJ 07871 (800) 227-0729 A/C Cooling Load calculation and measurement When we talk about sizing an air conditioning appliance (tons of cooling, BTU/h or KW), we are specifying

More information

Session: HVAC 101 HVAC 101. Steve Sain Sain Engineering Associates, Inc. August 9, Rhode Island Convention Center Providence, Rhode Island

Session: HVAC 101 HVAC 101. Steve Sain Sain Engineering Associates, Inc. August 9, Rhode Island Convention Center Providence, Rhode Island Session: HVAC 101 HVAC 101 Steve Sain Sain Engineering Associates, Inc. August 9, 2016 Rhode Island Convention Center Providence, Rhode Island Why? 2 Acknowledgements 3 Disclaimer I m gonna shoot down

More information

AHRI Standard 1250P (I-P) 2009 Standard for Performance Rating of Walk-In Coolers and Freezers

AHRI Standard 1250P (I-P) 2009 Standard for Performance Rating of Walk-In Coolers and Freezers AHRI Standard 1250P (I-P) 2009 Standard for Performance Rating of Walk-In Coolers and Freezers - ERRATA SHEET FOR AHRI STANDARD 1250 (I-P)-2009, PERFORMANCE RATING OF WALK-IN COOLERS AND FREEZERS December

More information

A Treatise on Liquid Subcooling

A Treatise on Liquid Subcooling A Treatise on Liquid Subcooling While the subject of this article is Liquid Refrigerant Subcooling, its affect on the operation of the thermostatic expansion valve (TEV), and ultimately on system performance

More information

ME 410 MECHA ICAL E GI EERI G SYSTEMS LABORATORY

ME 410 MECHA ICAL E GI EERI G SYSTEMS LABORATORY ME 410 MECHA ICAL E GI EERI G SYSTEMS LABORATORY MASS & E ERGY BALA CES I PSYCHROMETRIC PROCESSES EXPERIME T 3 1. OBJECTIVE The object of this experiment is to observe four basic psychrometric processes

More information

SYNOPSIS. Part-Load Control Strategies for Packaged Rooftop Units. In this issue... Bin Hour Profile Charlotte, NC

SYNOPSIS. Part-Load Control Strategies for Packaged Rooftop Units. In this issue... Bin Hour Profile Charlotte, NC VOLUME ONE NUMBER THREE SYNOPSIS A N H V A C N E W S L E T T E R F O R B U I L D I N G O W N E R S A N D M A N A G E R S In this issue... Part-Load Strategies Why they re important.......... 1 What Things

More information

Summary of Comments (Washington Revisions November 7, 2000) Update November 27, 2000

Summary of Comments (Washington Revisions November 7, 2000) Update November 27, 2000 SAE Alternate Refrigerant Cooperative Research Program Summary of Comments (Washington Revisions November 7, 2000) Update November 27, 2000 To: Alternate Refrigerant Task Force Members From: Ward Atkinson

More information

Air Conditioning Clinic

Air Conditioning Clinic Air Conditioning Clinic Psychrometry One of the Fundamental Series D C B A C B D A July 2012 TRG-TRC001-EN Psychrometry One of the Fundamental Series A publication of Trane Preface Psychrometry A Trane

More information

CHAPTER 7 PERFORMANCE ANALYSIS OF VAPOUR COMPRESSION REFRIGERATION SYSTEM IN HYBRID REFRIGERATION SYSTEM

CHAPTER 7 PERFORMANCE ANALYSIS OF VAPOUR COMPRESSION REFRIGERATION SYSTEM IN HYBRID REFRIGERATION SYSTEM 111 CHAPTER 7 PERFORMANCE ANALYSIS OF VAPOUR COMPRESSION REFRIGERATION SYSTEM IN HYBRID REFRIGERATION SYSTEM 7.1 INTRODUCTION Energy is the primary component to run any system in the world. According to

More information

Calculated Energy Savings for Disabling Anti- Sweat Door Heat on Glass Display Case Doors

Calculated Energy Savings for Disabling Anti- Sweat Door Heat on Glass Display Case Doors Calculated Energy Savings for Disabling Anti- Sweat Door Heat on Glass Display Case Doors ET Project Number: Product Manager: Sachin Andhare Pacific Gas and Electric Company` Prepared By: Danielle Geers

More information

LOW RECIRCULATION RATE EVAPORATORS

LOW RECIRCULATION RATE EVAPORATORS WHITE PAPER LOW RECIRCULATION RATE EVAPORATORS Written by: Jeff Welch President Welch Engineering Corporation Abstract Fin coil evaporators with enhanced internal tube surfaces allow for optimum performance

More information

Scientific Principals and Analytical Model. Charcoal Cooler. Lisa Crofoot MECH 425, Queens University

Scientific Principals and Analytical Model. Charcoal Cooler. Lisa Crofoot MECH 425, Queens University Scientific Principals and Analytical Model Charcoal Cooler Lisa Crofoot MECH 425, Queens University 1.0 Scientific Principles Evaporative cooling is based on the principle that water requires heat energy

More information

Small Commercial Business Energy Audits. Recognizing and addressing the special requirements of the small business market segment.

Small Commercial Business Energy Audits. Recognizing and addressing the special requirements of the small business market segment. Small Commercial Business Energy Audits Recognizing and addressing the special requirements of the small business market segment. Class 7 Class 7: Refrigeration Fundamentals The Vapor-compression cycle

More information

CEC Evaluation of Horizontal Recirculatory Air Curtain Efficiencies - Cooler to Conditioned Space.

CEC Evaluation of Horizontal Recirculatory Air Curtain Efficiencies - Cooler to Conditioned Space. ASME 2014 Citrus Engineering Conference CEC2014 June 12, 2014, Lake Alfred, Florida, USA CEC2014-5801 Evaluation of Horizontal Recirculatory Air Curtain Efficiencies - Cooler to Conditioned Space. Daniel

More information

Implementation and testing of a model for the calculation of equilibrium between components of a refrigeration installation

Implementation and testing of a model for the calculation of equilibrium between components of a refrigeration installation Implementation and testing of a model for the calculation of equilibrium between components of a refrigeration installation Marco Alexandre Onofre Dias This paper presents a model that was implemented

More information

97.501B. Prepared by: (415) Sponsored By: Final Report. Author: Presented at: ASHRAE. Meeting SF Me!

97.501B. Prepared by: (415) Sponsored By: Final Report. Author: Presented at: ASHRAE. Meeting SF Me! 97.501B Prepared by: Proctor Engineering Group, Ltd. San Rafael, CA 94901 (415) 451-24800 Monitored In-Situ Performance of Residential Air-Co onditioning Systems Sponsored By: Arizona Public Service Company

More information

9. ENERGY PERFORMANCE ASSESSMENT OF HVAC SYSTEMS

9. ENERGY PERFORMANCE ASSESSMENT OF HVAC SYSTEMS 9. ENERGY PERFORMANCE ASSESSMENT OF HVAC SYSTEMS 9.1 Introduction Air conditioning and refrigeration consume significant amount of energy in buildings and in process industries. The energy consumed in

More information

Performance of an Improved Household Refrigerator/Freezer

Performance of an Improved Household Refrigerator/Freezer Performance of an Improved Household Refrigerator/Freezer D. O. Ariyo Department of Mechanical Engineering, School of Engineering, Federal Polytechnic, Offa, Nigeria Y. L. Shuaib-Babata Department of Mechanical

More information

MODELLING AND OPTIMIZATION OF DIRECT EXPANSION AIR CONDITIONING SYSTEM FOR COMMERCIAL BUILDING ENERGY SAVING

MODELLING AND OPTIMIZATION OF DIRECT EXPANSION AIR CONDITIONING SYSTEM FOR COMMERCIAL BUILDING ENERGY SAVING MODELLING AND OPTIMIZATION OF DIRECT EXPANSION AIR CONDITIONING SYSTEM FOR COMMERCIAL BUILDING ENERGY SAVING V. Vakiloroaya*, J.G. Zhu, and Q.P. Ha School of Electrical, Mechanical and Mechatronic Systems,

More information

Working Draft Measure Information Template Refrigerated Warehouse

Working Draft Measure Information Template Refrigerated Warehouse Page 1 CODES AND STANDARDS ENHANCEMENT INITIATIVE (CASE) Working Draft Measure Information Template Refrigerated Warehouse 2013 California Building Energy Efficiency Standards California Utilities Statewide

More information

International Journal of Engineering & Technology Sciences Volume 03, Issue 01, Pages 55-64, 2015

International Journal of Engineering & Technology Sciences Volume 03, Issue 01, Pages 55-64, 2015 International Journal of Engineering & Technology Sciences Volume 03, Issue 01, Pages 55-64, 2015 ISSN: 2289-4152 Performance Evaluation of Vapour Compression Refrigeration System Using Double Effect Condensing

More information

2 Building Type. 3 Trade Ally/Contractor Information. Refrigeration Rebate Worksheet. Retrofit New Construction

2 Building Type. 3 Trade Ally/Contractor Information. Refrigeration Rebate Worksheet. Retrofit New Construction Refrigeration Rebate Worksheet Before you start Instructions: Complete all relevant information for your project. Include with completed final application package. All rebates are paid per unit and are

More information

COMPARING AIR COOLER RATINGS PART 1: Not All Rating Methods are Created Equal

COMPARING AIR COOLER RATINGS PART 1: Not All Rating Methods are Created Equal By Bruce I. Nelson, P.E., President, Colmac Coil Manufacturing, Inc. COMPARING AIR COOLER RATINGS PART 1: Not All Rating Methods are Created Equal Summary Refrigeration air coolers (evaporators) are widely

More information

Sustainable Techniques in Refrigerated Space

Sustainable Techniques in Refrigerated Space Sustainable Techniques in Refrigerated Space Sustainability and high performance of refrigerated space used for the preservation of perishable product capitalizes on conservation techniques which reduce

More information

Appendix A. Glossary of Common Terms

Appendix A. Glossary of Common Terms Glossary of Common Terms Glossary of Common Terms Absorption chiller A refrigeration machine using heat as the power input to generate chilled water. Adjustable speed drive A means of changing the speed

More information

5. ASSESSMENT RECOMMENDATIONS

5. ASSESSMENT RECOMMENDATIONS 5. ASSESSMENT RECOMMENDATIONS AR No. 1 Reduce Discharge Pressure Recommended Action Reduce the minimum condensing temperatures on your warehouse refrigeration systems from 110 F to 60 F. In order to reduce

More information

Analysis of Constant Pressure and Constant Area Mixing Ejector Expansion Refrigeration System using R-1270 as Refrigerant

Analysis of Constant Pressure and Constant Area Mixing Ejector Expansion Refrigeration System using R-1270 as Refrigerant Analysis of Constant Pressure and Constant Area Mixing Ejector Expansion Refrigeration System using R-1270 as Refrigerant Ravi Verma 1, Sharad Chaudhary 2 1, 2 Department of Mechanical Engineering, IET

More information

product application data PERFECT HUMIDITY DEHUMIDIFICATION SYSTEM

product application data PERFECT HUMIDITY DEHUMIDIFICATION SYSTEM product application data PERFECT HUMIDITY DEHUMIDIFICATION SYSTEM 551B 581B DuraPac Plus Series Sizes 036-150 3to12 1 / 2 Tons Cancels: New PAD 551B.36.2 10/1/04 CONTENTS Page INTRODUCTION.................................

More information

Computerized Simulation of Automotive Air-Conditioning System: A Parametric Study

Computerized Simulation of Automotive Air-Conditioning System: A Parametric Study www.ijcsi.org 787 Computerized Simulation of Automotive Air-Conditioning System: A Parametric Study Haslinda Mohamed Kamar 1, Nazri Kamsah 2 and Mohd Yusoff Senawi 3 1 Department of Thermo-Fluid, Universiti

More information

Evaluation of Low Pressure Dryer for Plastic Resins

Evaluation of Low Pressure Dryer for Plastic Resins Design & Engineering Services Evaluation of Low Pressure Dryer for Plastic Resins ET 08.09 Final Report Design & Engineering Services Customer Service Business Unit Southern California Edison July 27,

More information

SECTION 7 AIR CONDITIONING (COOLING) UNIT 40 TYPICAL OPERATING CONDITIONS

SECTION 7 AIR CONDITIONING (COOLING) UNIT 40 TYPICAL OPERATING CONDITIONS SECTION 7 AIR CONDITIONING (COOLING) UNIT 40 TYPICAL OPERATING CONDITIONS UNIT OBJECTIVES After studying this unit, the reader should be able to Explain what conditions will cause the evaporator pressure

More information

SECTION 7 AIR CONDITIONING (COOLING) UNIT 40 TYPICAL OPERATING CONDITIONS UNIT OBJECTIVES

SECTION 7 AIR CONDITIONING (COOLING) UNIT 40 TYPICAL OPERATING CONDITIONS UNIT OBJECTIVES SECTION 7 AIR CONDITIONING (COOLING) UNIT 40 TYPICAL OPERATING CONDITIONS UNIT OBJECTIVES After studying this unit, the reader should be able to Explain what conditions will cause the evaporator pressure

More information

AIR CONDITIONING. Carrier Corporation 2002 Cat. No

AIR CONDITIONING. Carrier Corporation 2002 Cat. No AIR CONDITIONING Carrier Corporation 2002 Cat. No. 020-016 1. This refresher course covers topics contained in the AIR CONDITIONING specialty section of the North American Technician Excellence (NATE)

More information

Development of a Psychrometric Test Chamber. Michael J. Swedish. Associate Professor Mechanical Engineering Department Milwaukee School of Engineering

Development of a Psychrometric Test Chamber. Michael J. Swedish. Associate Professor Mechanical Engineering Department Milwaukee School of Engineering Session 2633 Development of a Psychrometric Test Chamber Michael J. Swedish Associate Professor Mechanical Engineering Department Milwaukee School of Engineering Acknowledgments The design of the Psychrometric

More information

Applications of Thermodynamics: Heat Pumps and Refrigerators

Applications of Thermodynamics: Heat Pumps and Refrigerators Applications of Thermodynamics: Heat Pumps and Refrigerators Bởi: OpenStaxCollege Almost every home contains a refrigerator. Most people don t realize they are also sharing their homes with a heat pump.

More information

Low GWP Refrigerants for Air Conditioning Applications

Low GWP Refrigerants for Air Conditioning Applications Purdue University Purdue e-pubs International Refrigeration and Air Conditioning Conference School of Mechanical Engineering 2014 Low GWP Refrigerants for Air Conditioning Applications Samuel F. Yana Motta

More information

MODELING OF THE SINGLE COIL, TWIN FAN AIR-CONDITIONING SYSTEM IN ENERGYPLUS

MODELING OF THE SINGLE COIL, TWIN FAN AIR-CONDITIONING SYSTEM IN ENERGYPLUS MODELING OF THE SINGLE COIL, TWIN FAN AIR-CONDITIONING SYSTEM IN ENERGYPLUS Clayton Miller 1,* and Chandra Sekhar 1 1 National University of Singapore, Singapore * Corresponding email: miller.clayton@nus.edu.sg

More information

Thomas J Kelly. Fundamentals of Refrigeration. Sr. Engineering Instructor Carrier Corporation. August 20, Page number: 1.

Thomas J Kelly. Fundamentals of Refrigeration. Sr. Engineering Instructor Carrier Corporation. August 20, Page number: 1. Thomas J Kelly Sr. Engineering Instructor Carrier Corporation August 20, 2003 1 SESSION OBJECTIVES At the conclusion of this session you should be able to: 1. Describe the basics principles of refrigeration

More information

TEST REPORT #32. System Soft-Optimized Test of Refrigerant D2Y60 in Air Source Heat Pump

TEST REPORT #32. System Soft-Optimized Test of Refrigerant D2Y60 in Air Source Heat Pump Air-Conditioning, Heating, and Refrigeration Institute (AHRI) Low-GWP Alternative Refrigerants Evaluation Program (Low-GWP AREP) TEST REPORT #32 System Soft-Optimized Test of Refrigerant D2Y60 in Air Source

More information

Displacement Ventilation in Classrooms

Displacement Ventilation in Classrooms TWO DIFFUSERS, THE WHITE COLUMNS ON EACH SIDE OF THE WHITEBOARD, PROVIDE COOL SUPPLY AIR TO THE DEMONSTRATION DV CLASSROOM. THE DV CLASSROOM DEMONSTRATION FEATURED A CUSTOMIZED ROOFTOP HVAC UNIT. Displacement

More information

ENERGY SAVINGS THROUGH LIQUID PRESSURE AMPLIFICATION IN A DAIRY PLANT REFRIGERATION SYSTEM. A. Hadawey, Y. T. Ge, S. A. Tassou

ENERGY SAVINGS THROUGH LIQUID PRESSURE AMPLIFICATION IN A DAIRY PLANT REFRIGERATION SYSTEM. A. Hadawey, Y. T. Ge, S. A. Tassou Condensers ENERGY SAVINGS THROUGH LIQUID PRESSURE AMPLIFICATION IN A DAIRY PLANT REFRIGERATION SYSTEM A. Hadawey, Y. T. Ge, S. A. Tassou 4 7 8 5 3 2 6 1 Produced by: The Centre for Energy and Built Environment

More information

Some Modeling Improvements for Unitary Air Conditioners and Heat Pumps at Off-Design Conditions

Some Modeling Improvements for Unitary Air Conditioners and Heat Pumps at Off-Design Conditions Purdue University Purdue e-pubs International Refrigeration and Air Conditioning Conference School of Mechanical Engineering 2006 Some Modeling Improvements for Unitary Air Conditioners and Heat Pumps

More information

Comfort and health-indoor air quality

Comfort and health-indoor air quality Comfort and health-indoor air quality 1 The human body has a complicated regulating system to maintain the human body temperature constant most of the time, which is 98.6 F (36.9 C) regardless of the environmental

More information

Low Global Warming Refrigerants For Commercial Refrigeration Systems

Low Global Warming Refrigerants For Commercial Refrigeration Systems Purdue University Purdue e-pubs International Refrigeration and Air Conditioning Conference School of Mechanical Engineering 2012 Low Global Warming Refrigerants For Commercial Refrigeration Systems Samuel

More information

Application of Air Source Variable Refrigerant Flow in Cold Climates

Application of Air Source Variable Refrigerant Flow in Cold Climates PREPARED BY Seventhwave with the assistance of Daikin North America, LLC Masters Building Solutions Application of Air Source Variable Refrigerant Flow in Cold Climates A White Paper March 2015 275-1

More information

2013 Guideline for Specifying the Thermal Performance of Cool Storage Equipment. AHRI Guideline T (I-P)

2013 Guideline for Specifying the Thermal Performance of Cool Storage Equipment. AHRI Guideline T (I-P) 2013 Guideline for Specifying the Thermal Performance of Cool Storage Equipment AHRI Guideline T (I-P) IMPORTANT SAFETY DISCLAIMER AHRI does not set safety standards and does not certify or guarantee the

More information

AR/IA/UP 241 Lecture 5: Psychrometrics

AR/IA/UP 241 Lecture 5: Psychrometrics Faculty of Architecture and Planning Thammasat University AR/IA/UP 241 Lecture 5: Psychrometrics Author: Asst. Prof. Chalermwat Tantasavasdi 1. Definition of Psychrometric Chart The word psychrometry is

More information

OFF-GRID COMMERCIAL DIRECT CURRENT GRID SYSTEM

OFF-GRID COMMERCIAL DIRECT CURRENT GRID SYSTEM Design & Engineering Services OFF-GRID COMMERCIAL DIRECT CURRENT GRID SYSTEM Report Prepared by: Design & Engineering Services Customer Service Business Unit Southern California Edison December 2012 Acknowledgements

More information

For an ideal gas mixture, Dalton s law states that the sum of the partial pressures of the individual components is equal to the total pressure.

For an ideal gas mixture, Dalton s law states that the sum of the partial pressures of the individual components is equal to the total pressure. 1 PSYCHROMETICS Psychrometry is the study of the characteristics of moist air. We will see soon that evaporation of moisture from the skin can have a significant impact on thermal comfort. The rate of

More information

Fan Selection. and Energy Savings

Fan Selection. and Energy Savings Fan Selection and Energy Savings Reducing power consumption by minimizing system pressure loss and choosing the right equipment HBy BRIAN MLEZIVA Greenheck Fan Corp. Schofield, Wis. Power Consumption HVAC

More information

DEMONSTRATION OF A MICROCHANNEL HEAT EXCHANGER FOR OPERATION IN A REVERSIBLE HEAT PUMP SYSTEM

DEMONSTRATION OF A MICROCHANNEL HEAT EXCHANGER FOR OPERATION IN A REVERSIBLE HEAT PUMP SYSTEM DEMONSTRATION OF A MICROCHANNEL HEAT EXCHANGER FOR OPERATION IN A REVERSIBLE HEAT PUMP SYSTEM Hantz, D., Gulyas, G., Bensafi, A. Centre Technique des Industries Aéraulique et Thermique (CETIAT) BP 2042

More information

System Modeling of Gas Engine Driven Heat Pump

System Modeling of Gas Engine Driven Heat Pump Purdue University Purdue e-pubs International Refrigeration and Air Conditioning Conference School of Mechanical Engineering 01 System Modeling of Gas Engine Driven Heat Pump Isaac Mahderekal shenb@ornl.gov

More information

Cold Storage Warehouse Dock Parametric Study

Cold Storage Warehouse Dock Parametric Study Cold Storage Warehouse Dock Parametric Study Todd B. Jekel, Ph.D. Industrial Refrigeration Consortium Topics Objective Modeling loads Infiltration Desiccant operation Effects of dock temperature setpoint

More information

Design of Divided Condensers for Desiccant Wheel-Assisted Separate Sensible and Latent Cooling AC Systems

Design of Divided Condensers for Desiccant Wheel-Assisted Separate Sensible and Latent Cooling AC Systems Purdue University Purdue e-pubs International Refrigeration and Air Conditioning Conference School of Mechanical Engineering 2012 Design of Divided Condensers for Desiccant Wheel-Assisted Separate Sensible

More information

TEST REPORT #4. Travis Crawford Dutch Uselton. Lennox Industries Inc Metrocrest Drive Carrollton, TX 75006

TEST REPORT #4. Travis Crawford Dutch Uselton. Lennox Industries Inc Metrocrest Drive Carrollton, TX 75006 Air-Conditioning, Heating, and Refrigeration Institute (AHRI) Low-GWP Alternative Refrigerants Evaluation Program (Low-GWP AREP) TEST REPORT #4 System Drop-in Test of Refrigerant R-32 in Split System Heat

More information

Technical Development Program. COMMERCIAL HVAC PACKAGED EQUIPMENT Split Systems PRESENTED BY: Ray Chow Sigler

Technical Development Program. COMMERCIAL HVAC PACKAGED EQUIPMENT Split Systems PRESENTED BY: Ray Chow Sigler Technical Development Program COMMERCIAL HVAC PACKAGED EQUIPMENT Split Systems PRESENTED BY: Ray Chow Sigler Menu Section 1 Section 2 Section 3 Section 4 Section 5 Section 6 Section 7 Introduction System

More information

Experimental Study on Match for Indoor and Outdoor Heat Exchanger of Residential Airconditioner

Experimental Study on Match for Indoor and Outdoor Heat Exchanger of Residential Airconditioner Purdue University Purdue e-pubs International Refrigeration and Air Conditioning Conference School of Mechanical Engineering 2014 Experimental Study on Match for Indoor and Outdoor Heat Exchanger of Residential

More information

Energy Savings Potential of Passive Chilled Beam System as a Retrofit Option for Commercial Buildings in Different Climates

Energy Savings Potential of Passive Chilled Beam System as a Retrofit Option for Commercial Buildings in Different Climates Purdue University Purdue e-pubs International High Performance Buildings Conference School of Mechanical Engineering 2014 Energy Savings Potential of Passive Chilled Beam System as a Retrofit Option for

More information

Feasibility of a Liquid Desiccant Application in an Evaporative. Cooling Assisted 100% Outdoor Air System

Feasibility of a Liquid Desiccant Application in an Evaporative. Cooling Assisted 100% Outdoor Air System Feasibility of a Liquid Desiccant Application in an Evaporative Cooling Assisted 100% Outdoor Air System M.H.Kim 1, S.K.Han 1, S.Y. Cho 2, and J.W.Jeong 1* 1 Department of Architectural Engineering, Hanyang

More information

CTI Sponsored Educational Program

CTI Sponsored Educational Program Presented By: Kent Martens SPX Cooling Technologies, Inc. Slide No.: 1 CTI Mission Statement To advocate and promote the use of environmentally responsible Evaporative Heat Transfer Systems (EHTS) for

More information

CHAPTER 2 EXPERIMENTAL APPARATUS AND PROCEDURES

CHAPTER 2 EXPERIMENTAL APPARATUS AND PROCEDURES CHAPTER 2 EXPERIMENTAL APPARATUS AND PROCEDURES The experimental system established in the present study to investigate the transient flow boiling heat transfer and associated bubble characteristics of

More information

March 2014 AHRI Presents Research Findings at Low-GWP AREP Conference

March 2014 AHRI Presents Research Findings at Low-GWP AREP Conference March 2014 AHRI Presents Research Findings at Low-GWP AREP Conference In addition to its regular research activities, AHRI is conducting the Low-Global Warming Potential Alternative Refrigerants Evaluation

More information

Analysis of freeze protection methods for recuperators used in energy recovery from exhaust air

Analysis of freeze protection methods for recuperators used in energy recovery from exhaust air Analysis of freeze protection methods for recuperators used in energy recovery from exhaust air Anna Pacak 1,*, Andrzej Jedlikowski 1, Demis Pandelidis 1, and Sergey Anisimov 1 1 Wrocław University of

More information

Drying principles and general considerations

Drying principles and general considerations Drying principles and general considerations Drying Mechanisms In the process of drying heat is necessary to evaporate moisture from the grain and a flow of air is needed to carry away the evaporated moisture.

More information

S.A. Klein and G.F. Nellis Cambridge University Press, 2011

S.A. Klein and G.F. Nellis Cambridge University Press, 2011 12.A-1 A mixture of helium and water vapor is flowing through a pipe at T= 90 C and P = 150 kpa. The mole fraction of helium is y He = 0.80. a.) What is the relative humidity of the mixture? b.) What is

More information

Refrigerant Mass and Oil Migration During Start-up Transient

Refrigerant Mass and Oil Migration During Start-up Transient Refrigerant Mass and Oil Migration During Start-up Transient Steffen Peuker, Predrag Hrnjak* Department of Mechanical and Industrial Engineering University of Illinois at Urbana-Champaign 1206 West Green

More information

Math. The latent heat of fusion for water is 144 BTU s Per Lb. The latent heat of vaporization for water is 970 Btu s per Lb.

Math. The latent heat of fusion for water is 144 BTU s Per Lb. The latent heat of vaporization for water is 970 Btu s per Lb. HVAC Math The latent heat of fusion for water is 144 BTU s Per Lb. The latent heat of vaporization for water is 970 Btu s per Lb. Math F. to C. Conversion = (f-32)*(5/9) C. to F. Conversion = C * 9/5 +

More information

Redesign of Bennett Hall HVAC System

Redesign of Bennett Hall HVAC System MEE 488 April 18, 2006 Redesign of Bennett Hall HVAC System Greg Andreasen Michael Chicoine Florent Hohxa Jason Jacobe Mechanical Engineering, University of Maine, Orono ME 04473, USA ABSTRACT Our task

More information

Don t Turn Active Beams Into Expensive Diffusers

Don t Turn Active Beams Into Expensive Diffusers This article was published in ASHRAE Journal, April 2012. Copyright 2012 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. Posted at www.ashrae.org. This article may not be

More information

Measure Guidelines. for EnergySmart Grocer (ESG)

Measure Guidelines. for EnergySmart Grocer (ESG) Measure Guidelines for EnergySmart Grocer (ESG) Measure Guidelines for EnergySmart Grocer (ESG) Table of Contents Page Cases 1. Low Open Case to New Reach-in..................................1 2. Medium

More information

Effects of Frost Formation on the External Heat Transfer Coefficient of a Counter-Crossflow Display Case Air Coil

Effects of Frost Formation on the External Heat Transfer Coefficient of a Counter-Crossflow Display Case Air Coil Purdue University Purdue e-pubs International Refrigeration and Air Conditioning Conference School of Mechanical Engineering 1998 Effects of Frost Formation on the External Heat Transfer Coefficient of

More information

PREDICTION OF THE PRESSURE DROP OF CO 2 IN AN EVAPORATOR USED FOR AIR COOLING ABSTRACT 1. INTRODUCTION 2. EXPERIMENTAL SET-UP AND PROCEDURE

PREDICTION OF THE PRESSURE DROP OF CO 2 IN AN EVAPORATOR USED FOR AIR COOLING ABSTRACT 1. INTRODUCTION 2. EXPERIMENTAL SET-UP AND PROCEDURE PREDICTION OF THE PRESSURE DROP OF CO 2 IN AN EVAPORATOR USED FOR AIR COOLING M. POIRIER, D. GIGUÈRE, Z. AIDOUN, M. OUZZANE Natural Resources Canada, CANMET Energy Technology Centre-Varennes 1615, Lionel

More information

Study on the Effect of Blade Angle on the Performance of a Small Cooling Tower

Study on the Effect of Blade Angle on the Performance of a Small Cooling Tower Kasetsart J. (Nat. Sci.) 42 : 378-384 (2008) Study on the Effect of Blade Angle on the Performance of a Small Cooling Tower Montri Pirunkaset* and Santi Laksitanonta ABSTRACT This paper presents the effect

More information

Maluna Unhinged vs Yeti Tundra Thermal Performance Comparison January 2017

Maluna Unhinged vs Yeti Tundra Thermal Performance Comparison January 2017 Maluna Unhinged vs Yeti Tundra Thermal Performance Comparison January 2017 Abstract A Maluna Unhinged cooler was tested and measured for performance relative to a Yeti Tundra cooler. Three identical tests

More information

Application of two hybrid control methods of expansion valves and vapor injected compression to heat pumps

Application of two hybrid control methods of expansion valves and vapor injected compression to heat pumps AM-4249223-1 - Application of two hybrid control methods of expansion valves and vapor injected compression to heat pumps Christian K. Bach, Graduate Student, Eckhard A. Groll, Reilly Professor, James

More information

Adding More Fan Power Can Be a Good Thing

Adding More Fan Power Can Be a Good Thing This article was published in ASHE Journal, May 14. Copyright 14 ASHE. Posted at www.ashrae.org. This article may not be copied and/or distributed electronically or in paper form without permission of

More information

Psychrometric Engineering Applications

Psychrometric Engineering Applications Psychrometric Engineering Applications Course No: M01-014 Credit: 1 PDH Klas Haglid, P.E., R.A. Continuing Education and Development, Inc. 9 Greyridge Farm Court Stony Point, NY 10980 P: (877) 322-5800

More information

RSES Technical Institute Training Manual 2 72 hours, 72 NATE CEHs, 7.2 CEUs

RSES Technical Institute Training Manual 2 72 hours, 72 NATE CEHs, 7.2 CEUs Lesson 1 - Trade Tools Explain the importance of using proper tools and test instruments. List the various types of wrenches and describe their use. Describe the proper procedures for bending, flaring,

More information

Psychrometrics. Outline. Psychrometrics. What is psychrometrics? Psychrometrics in daily life and food industry Psychrometric chart

Psychrometrics. Outline. Psychrometrics. What is psychrometrics? Psychrometrics in daily life and food industry Psychrometric chart Psychrometrics Outline What is psychrometrics? Psychrometrics in daily life and food industry Psychrometric chart Dry bulb temperature, wet bulb temperature, absolute humidity, relative humidity, specific

More information

Technical Papers. 32nd Annual Meeting. International Institute of Ammonia Refrigeration. March 14 17, 2010

Technical Papers. 32nd Annual Meeting. International Institute of Ammonia Refrigeration. March 14 17, 2010 Technical Papers 32nd Annual Meeting International Institute of Ammonia Refrigeration March 14 17, 2010 2010 Industrial Refrigeration Conference & Exhibition Manchester Grand Hyatt San Diego, California

More information

A Generalized Correlation for Pressure Drop of Refrigerant R-134a through Helical Capillary Tubes

A Generalized Correlation for Pressure Drop of Refrigerant R-134a through Helical Capillary Tubes ISSN 2395-1621 A Generalized Correlation for Pressure Drop of Refrigerant R-134a through Helical Capillary Tubes #1 N. Dhekale, #2 Dr. P. A. Patil 1 amardhekale@gmail.com #12 Research Laboratory, Department

More information

Fans: Features and Analysis

Fans: Features and Analysis Technical Development Program COMMERCIAL HVAC EQUIPMENT Fans: Features and Analysis PRESENTED BY: Michael Ho Version 1.2 Menu Section 1 Section 2 Section 3 Section 4 Section 5 Section 6 Section 7 Section

More information

TESTS OF ADSIL COATING

TESTS OF ADSIL COATING TESTS OF ADSIL COATING Test B - Long Term Test FSEC-CR-1259-01 July 11, 2001 Prepared for: Bob Suggs Florida Power & Light Company 9250 W. Flagler Street Miami, Florida 33174 Principal Investigator Dr.

More information

The Electronic Newsletter of The Industrial Refrigeration Consortium Vol. 11 No. 1, 2011

The Electronic Newsletter of The Industrial Refrigeration Consortium Vol. 11 No. 1, 2011 chan The Electronic Newsletter of The Industrial Refrigeration Consortium Vol. 11 No. 1, 2011 PRESSURE RELIEF CAPACITY DETERMINATION FOR EQUIPMENT Previously in The Cold Front (2006), we provided an introduction

More information

How to ensure Thermal Comfort in buildings with CFD

How to ensure Thermal Comfort in buildings with CFD How to ensure Thermal Comfort in buildings with CFD Using CFD Simulations to analyze and improve HVAC Systems Simulation in your browser How to Ensure Thermal Comfort in Buildings with CFD How to use CFD

More information

Pressure drop analysis of evaporator using refrigerants R-22, R-404A and R-407C

Pressure drop analysis of evaporator using refrigerants R-22, R-404A and R-407C drop analysis of evaporator using refrigerants R-22, R-404A and R-407C #1 Pallavi Sawkar, #2 Dr. Pradeep Patil #12 Department of Mechanical Engineering, JSPM s JSCOE, Savitribai Phule Pune University,

More information

INTEGRATION OF DEMAND RESPONSE INTO TITLE 24 FOR COMMERCIAL REFRIGERATION SYSTEMS

INTEGRATION OF DEMAND RESPONSE INTO TITLE 24 FOR COMMERCIAL REFRIGERATION SYSTEMS Design & Engineering Services INTEGRATION OF DEMAND RESPONSE INTO TITLE 24 FOR COMMERCIAL REFRIGERATION SYSTEMS Phase1: Demand Response Potential DR 09.06.01 Report Prepared by: Design & Engineering Services

More information

Performance Rating of Commercial and Industrial Unitary Air-Conditioning Condensing Units

Performance Rating of Commercial and Industrial Unitary Air-Conditioning Condensing Units ANSI/AHRI Standard 366 (SI) 2009 Standard for Performance Rating of Commercial and Industrial Unitary Air-Conditioning Condensing Units Approved by ANSI on 4 January 2011 IMPORTANT SAFETY DISCLAIMER AHRI

More information