HEAT RECOVERY DESIGN FOR STARBUCKS STORE RETROFITLABORATORY BENCH TEST April, 2016

Prepared By:Dr. Thomas BradleyAssociate ProfessorCollege of EngineeringColorado State University

Chris AndersonGraduate Research AssistantCollege of EngineeringColorado State University

Becca StockUndergraduate Research AssistantCollege of EngineeringColorado State University

Stephanie BarrProject ManagerInstitute for the Built EnvironmentColorado State University

Prepared For:Urano RobinsonDirector – Global Innovation & Technology – Global R&D

Starbucks Coffee

CONTENTSExecutive Summary.....................................................................................................................................3

Introduction.................................................................................................................................................4

Testing Model Assumptions........................................................................................................................4

Air Conditioning System Savings..............................................................................................................5

Total Waste Heat.................................................................................................................................5

Total Air Conditioning Hours...............................................................................................................6

Hot Water Preheat..................................................................................................................................7

Refrigerator Efficiency.............................................................................................................................8

Part 1: Laboratory vs. In-Store Data Logging.......................................................................................8

Part 2: Refrigeration Efficiency as a Function of Water Loop Temperature.........................................9

Part 3: Water-Cooled vs. Fan-Cooled.................................................................................................10

System Energy Use................................................................................................................................12

Pump.................................................................................................................................................13

Fan.....................................................................................................................................................13

Costs......................................................................................................................................................14

Updated Model.........................................................................................................................................14

Sensitivity Analysis.....................................................................................................................................15

Next steps..................................................................................................................................................16

Areas for further research.........................................................................................................................16

Appendix...................................................................................................................................................17

Appendix A............................................................................................................................................17

Building Energy Model.......................................................................................................................17

Appendix B............................................................................................................................................22

Heat Rejection Calculations...............................................................................................................22

Appendix C............................................................................................................................................23

Fan Duty Cycle...................................................................................................................................23

Appendix D............................................................................................................................................26

Refrigeration Efficiency......................................................................................................................26

Small Refrigeration Units (1 door refrigerator, 1 door freezer, 2 door refrigerator).........................26

Large Refrigeration Units (vertical food case, horizontal food case, ice machine)............................28

Refrigeration Load vs. Power Consumption.......................................................................................30

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Appendix E.............................................................................................................................................31

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EXECUTIVE SUMMARYIn accordance with task order #3 of the master research agreement between Colorado State University and Starbucks Coffee Company, the team from Colorado State University was responsible for performing a heat recovery and mechanical efficiency feasibility assessment. This assessment was completed on December 15, 2015. As part of the feasibility assessment, an addendum to the master research agreement was included. This addendum outlined a plan for preliminary laboratory testing of the water loop heat recovery system prior to the installation of the final design in store #14944.

The objective of this report is to document the results of this laboratory testing and to modify the system performance model accordingly. The following figure provides a side-by-side comparison of the values detailed in this report and the original modeled values:

Reduced Air Conditioning Load

13,050 kWh $2,870

Increased Heating Load

-380 therms -$400

Hot Water Preheat 9,100 kWh $1,980Refrigeration

Effi ciency-3,400 kWh -$750

System Energy Use-4,660 kWh -$1,030

EquipmentLabor

System Costs

Simple Payback2.8 years

Annual System Energy Savings

$4,200$3,500

Reduced Air Conditioning Load

15,000 kWh $3,300

Increased Heating Load

- -

Hot Water Preheat 8,140 kWh $1,790Refrigeration

Effi ciency2,700 kWh $600

System Energy Use -4,660 kWh -$1,030

EquipmentLabor

1.8 years

Annual System Energy Savings

System Costs$4,800$4,000

Simple Payback

Figure 1: Results from Laboratory Testing, Original Modeled Values

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Laboratory Tested Values Original Model Values

INTRODUCTIONThe first design report submitted by the team from CSU outlined four original design options, and provided a detailed feasibility analysis for the design option selected. The final design option selected was the water loop heat recovery system. As per the master research and development agreement between CSU and Starbucks Coffee Company, the final design will be installed in Starbucks store #14944, located at 1708 S. College Avenue in Fort Collins, CO.

As an addendum to the master research and development agreement, it was determined that prior to the actual in-store installation the system should be tested in a laboratory to minimize financial risks associated with uncertainties surrounding the system’s performance. These uncertainties included: validation of water loop effectiveness providing appropriate condenser cooling; heat output of each refrigeration unit at different operating conditions; refrigeration efficiency variation due to changes in environmental conditions; and proposed system energy consumption. This report outlines the results of the laboratory testing directed towards addressing each of these uncertainties.

TESTING MODEL ASSUMPTIONSA laboratory mockup of the proposed heat recovery system was designed to test as many components of the system as possible in the lab. A schematic of the system is shown in Figure 2.

Figure 2: Lab Set-up Schematic

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The refrigeration equipment used in the lab set up was provided by Starbucks in the standard air-cooled condition with the exception of the ice machine. A licensed refrigeration contractor was used to install the water cooled condensers on each of the units. The condensing units were then connected by ¾” PEX water line. PVC “T” fittings were used to house water temperature loggers. These loggers could be placed at any point in the cycle depending on the requirements of each individual test. The loop is controlled by a temperature sensor which cycles the fan to maintain a steady water loop temperature. There is also a deaerator tank just before the pump, which is designed to prevent air bubbles in the system. This tank also allows water to be added to fill the system as there is no water connection in the lab. In the sections below, details of each test run and the results of each are outlined below by model component.

Air Conditioning System SavingsThere are two factors that contribute to the air conditioning savings of the system: the amount of waste heat removed by the system and the number of hours per year that the air conditioning system is running.

Total Waste HeatTo understand how much waste heat each of the refrigeration units was producing, the water temperature before and after the condenser was measured. This allowed the total heat produced by the unit to be calculated as outlined in Appendix B. The heat produced by each unit is shown in Table 1.

Table 1: Waste Heat Produced by Unit

Refrigeration UnitWaste Heat (Btu/h)

1-Door Refrigerator 5902-Door Refrigerator 1,0601-Door Freezer 1,080Vertical Food Case 7,040Horizontal Food Case 3,390Ice Machine 8,620Total 21,780

To confirm this data, the total heat load of the system was measured during the test run. A sample of the data is shown in Figure 3.

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1:22:31 PM1:47:01 PM 2:11:31 PM 2:36:01 PM3:00:31 PM 3:25:01 PM 3:49:31 PM0

5,000

10,000

15,000

20,000

25,000

30,000

35,000

System Heat OutputBt

u/h

Figure 3: Total Waste Heat from Refrigeration Units

By comparing the data derived from these two methods, the total waste heat can be confirmed to be about 22,000 btu per hour.

Total Air Conditioning HoursSince the laboratory set up couldn’t be used to determine to number of hours per year that the air conditioning system is running, a simple building energy model was built. In this model internal heat gains from equipment and people were compared to estimated heat gains or losses through ventilation and the building walls and roof. If the total heat gains were larger than the losses, the air conditioning system would be required to remove the excess heat. For details on the building energy model see Appendix A. The results of this model for a store with the same layout as the model store but located in San Diego, CA are shown in Figure 4.

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Figure 4: Building Heat Loss/Gain

From this histogram, we can estimate that the number of hours the air conditioning system runs for approximately 7,291 hours per year or about 10 months per year. For the remaining 1,469 hours per year, the building is in heating mode. This means that the system actually causes the heating system to use more energy since it is not being assisted by the waste heat in the building.

Using the waste heat values from our lab testing and the building system model, the updated estimate of the air conditioning savings is as follows:

Air Conditioning Savings: 13,050 kWh per year

Additional Heating Costs: 380 therms per year

Hot Water Preheat In the data presented in December, the water use at the model store had been logged and used with the efficiency data of the heat exchanger to calculate the expected heat recovery potential. These calculations used a water loop temperature of 100°F. During laboratory testing it was determined that the refrigerators stop operating properly with water loop temperatures above 100°F, so the heat recovery calculations were repeated with the three loop test temperatures: 75°F, 85°F, and 95°F. Due to concerns about the size of the indirect hot water heater for the domestic hot water preheat, the heat recovery of the plate-and-frame heat exchanger was calculated for this application as well. The results are outlined in Figure 5.

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75 77 79 81 83 85 87 89 91 93 950

1,000

2,000

3,000

4,000

5,000

6,000

DHW with Indirect Water Heater

DHW with Plate and Frame Heat Ex-changer

Coffee Water

Water Loop temperature (°F)

Ener

gy S

avin

gs (k

Wh/

yr)

Figure 5: Preheat Savings by Loop Temperature

From these calculations, it was determined that the indirect hot water heater can recover about 13% more energy when preheating water for the domestic hot water heater than the plate-and-frame heat exchanger. Given the space considerations and considering that the indirect hot water heater is a larger investment than the plate-and-frame heat exchanger, it is recommended that the plate-and-frame heat exchanger be used for the domestic hot water preheat as well as the coffee water preheat. By comparing these calculations to the refrigeration efficiency data it was determined that 95°F is the optimal loop temperature to maximize savings. This means that the estimated total hot water preheat savings is as follows:

Total Water Preheat Savings: 8,250 kWh per year

Refrigerator EfficiencyOne of the main objectives of the laboratory testing was to analyze the effect of the water-cooled condensers on the refrigeration efficiency for each unit. This study can be divided into three main components: comparing the laboratory operation to the in-store operation, the refrigeration efficiency changes at varying water loop temperatures, and comparing the air-cooled operation to the water-cooled operation.

Part 1: Laboratory vs. In-Store Data LoggingThe goal of the first part of this study was to determine a relationship between laboratory data and real world data. Due to space constraints and issues regarding the disruption of normal store operation, only the three smaller, True Refrigerators were data logged in the store. The power draw of the equivalent refrigeration units were also data logged in the lab. A comparison between the in-store data and the laboratory data for these refrigerators is shown below:

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Table 2: In-Store Power Draw vs. Laboratory Power Draw

UnitIn-Store Average Power Draw, W

Laboratory Average Power Draw, W

1-Door Refrigerator 126.3 81.62-Door Refrigerator 460.2 182.3

1-Door Freezer 268.5 194

It is evident that the in-store refrigerators operate at a much higher average power than the same units in the lab. There are several possible reasons for this discrepancy. The first possibility is that the in-store refrigerators see a greater heat load than the laboratory refrigerators. The constant opening and closing of the doors allows heat in, and ultimately results in the compressor working harder and longer. There is not an additional heat load due to the food stored in the refrigerators because it is all delivered to the store cold. To account for the heat load discrepancy, the 14944 Starbucks store was monitored over the course of a week, and the number of times each refrigerator was opened was counted. Then, the refrigerators in the laboratory were data logged again, but this including periodic openings of the doors. The results are shown below:

Table 3: Laboratory Testing, Simulated Store Operations

UnitNumber of

Openings per HourResulting Laboratory

Average Power Draw, W1-Door Refrigerator 0.33 752-Door Refrigerator 4.6 177

1-Door Freezer 1.8 209

This study demonstrated that load conditions cannot account for the discrepancy between refrigerator operation in the lab and in the store. Another possible cause is the difference in the ambient air condition. The lab is kept at 70°F and the refrigerators are located in the center of a room where they are well ventilated. In the store, the room temperature is set at 68°F, but the refrigerators are located against a wall. It is likely that the air surrounding the condensers is poorly circulated and therefore significantly warmer than the room temperature. In addition, the in-store refrigerators are older and may be suffering from wear-and-tear as compared to the newer lab units. In store testing will be required to confirm refrigeration efficiency changes under normal operating conditions.

Part 2: Refrigeration Efficiency as a Function of Water Loop TemperatureThe second part of this study focused on understanding the effect of the water loop temperature on the efficiency of the refrigeration units. Several experiments were conducted which involved operating each refrigerator and data logging the entering water temperature for the condenser as well as the refrigerator power draw each second over a period of several hours. The temperature was incrementally varied to determine the refrigerator’s response to different water loop temperatures. The results are shown in below:

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Figure 6: Refrigerator Power Consumption as a Function of Entering Water Temperature

For the three smaller refrigerators, each data point represents an averaged value over one cycle of refrigeration. Using the data points at each of the three temperature increments, an equation was developed for the average power of each unit as a function of the entering water temperature to be used for modeling purposes.

For the two food cases, each point represents an averaged value over a period of about three hours of operation at a set water temperature. For the vertical food case, the compressor was loaded during the entire testing interval for all tests of the water-cooled condenser. This is likely an indication that the condenser coil is undersized.

The ice machine was treated separately. According to manufacturer specifications, an ice machine operating with a water-cooled condenser requires 4.3 kilowatt hours per 100 pounds of ice produced, or 0.043 kWh/lb of ice. Two water loop temperature increments were tested in lab, and in both instances, the ice machine maintained an energy consumption within 1% of manufacturer specifications. Therefore, it was determined that water loop temperature had negligible effect on total power consumption, but it was noted that there was some evidence of increased batch times, up to 15% longer, with higher loop temperatures.

Part 3: Water-Cooled vs. Fan-CooledPrior to installing the water-cooled condensers, all refrigeration units were first tested with their standard air-cooled condensers (excluding the ice machine, which was shipped with a water-cooled condenser).

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During laboratory testing described in part 2, it was observed that both loaded power and duty cycle were affected by changes in water loop temperature. Considering the heat output of each unit and the work requirement of the radiator, the operating point for the water loop temperature to maximize energy savings is 95°F. A comparison of the average power consumption of each unit is shown below:

Table 4: Lab Testing, Air-Cooled at 70°F vs. Water-Cooled at 95°F

UnitAir-Cooled Power

Draw, WWater-Cooled Power

Draw, W1-Door Refrigerator 81.6 89.12-Door Refrigerator 182.3 223.9

1-Door Freezer 194 230.1Vertical Food Case 884.5 1,359

Horizontal Food Case 641 692Ice Machine* 1,290 1,070

Total 3,273 3,664

*The air-cooled data for the ice machine is taken from manufacturer specifications. The ice machine cooled by a water-cooled condenser was tested and compared to manufacturer data, and energy consumption per pound of ice was within 1% of the quoted specs.

In most instances, the water-cooled case requires slightly greater average power than the air-cooled case, with the exception of the ice machine. The vertical food case, however, sees a significant increase in average power, almost 475 Watts. After analyzing the data, in the air-cooled case, the unit’s compressor cycled as expected. But in the water-cooled case, the unit’s compressor operated constantly.

This dramatic reduction in efficiency most likely suggests the need for a larger water-cooled condenser coil. As an example case to demonstrate the significance of a properly sized condenser coil versus an improperly sized condenser coil, the 1 door freezer was initially retrofitted with a 1/3 hp coaxial coil. However, after analyzing the power logged data of the unit, it was determined that the freezer experienced a drastic reduction of efficiency with increasing loop temperatures. The 1/3 hp coaxial coil was replaced with a 1 hp coaxial coil. The following figure illustrates the results. Due to time constraints, additional retrofits were not possible.

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Figure 7: Freezer Duty Cycle for Two Distinct Condenser Coil Sizes

From Figure 7, it is evident that the duty cycle of the unit is significantly impacted by the size of the condenser coil. For a water loop temperature of 98°F, the freezer duty cycle using an undersized condenser coil is about 75%, as opposed to about 53% for a properly sized coil.

A properly sized water-cooled condenser coil would likely reduce the modeled average power consumption of the vertical food case from 1,359 Watts to around 884 Watts, a reduction of 475 Watts. This equates to a difference in annual energy consumption of 4,160 kWh. However, because the lower number has not been experimentally validated, the model reflects the higher value for average power.

There is an additional caveat to the sharp reduction in the vertical food case efficiency. The air-cooled vertical food case tested in lab demonstrated multiple on/off refrigeration cycles, but after the water-cooled condenser was retrofit the compressor ran consistently. After inspection, it was discovered that the vertical food case in the 14944 store does not exhibit any compressor cycling, and instead runs constantly. It is likely that the average power consumption of the lab tested vertical unit is significantly lower than the average power of the same unit operating in the store. This discrepancy has significant implications for the model developed to predict the annual savings of the system.

It is likely that the vertical unit tested in the lab with a water-cooled condenser, which seems to be operating inefficiently relative to the air-cooled case, is actually operating similar to the air-cooled case in the store. If this is the case, it would negate the negative effects of refrigeration efficiency currently built into the model, a positive change of around 3,425 kWh per year.

Refrigeration Efficiency Decrease: 3,425 kWh per year

System Energy Use To find the net energy savings of the system, the total energy used to run the system needed to be determined. The laboratory tests to determine the electrical use for the circulation pump and the exterior radiator fans are outlined below.

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PumpDue to the large pressure drop of the water loop, a positive displacement pump was chosen to provide the necessary flow rate and pressure. A diagram of a positive displacement pump is shown in Figure 8. In this style of pump, the water coming in fills up the compartment and then the rotation of the pump moves that waster to the outlet side. Due to the design of this pump the flow rate is nearly independent of the pressure being produced by the pump. This is because a constant volume of water is being moved from the inlet to the outlet for every rotation of the motor independent of any pressure changes. On the pump curve for the chosen pump, the flow rate varies from 5.43 to 5.51 gallons-per-minute over a pressure change of 50 to 250 psi. The only difference in operation at the higher operating pressures is the higher pressures require a larger input of energy from the motor.

In the lab, the flow rate of the system was measured to be 5.5 gallons-per-minute and the power of the pump was logged to be 400-watts. The pump runs consistently so this can be used to find the annual energy use of the pump.

Pump Energy Use: 3,500 kWh per year

FanTo reject the waste heat outside, two radiator and fan pairs are used. The fan controls are connected to a temperature sensor that cycle the fan on or off to maintain a constant water temperature. The first step to determining the amount of time each fan will run was to determine the heat transfer rate of the radiator with the fan on and with the fan off. To do this cold water was run thought the radiator and the temperature of the water was recorded before and after the radiator with the fan on and fan off. Using these heat rejection efficiency values and the total measured waste heat from the refrigeration units, the fan duty cycle could be calculated as a function of outside air temperature. For details of this experiment and calculations see Appendix C. The fan duty cycle by outside air temperature along with the number of hours at that temperature are shown in Figure 9.

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Figure 8: Positive Displacement Pump

0%

20%

40%

60%

80%

100%

120%

0

2

4

6

8

10

12

Outside Air Temperature Fan #1Fan #2

Outside Air Temperature (F)

Fan

Duty

Cyc

le

Hour

s Per

Yea

r

Figure 9: Fan Run Time Data

By multiplying the number of hours at each temperature with the duty cycle at that temperature, the number of hours the fan will run per year can be calculated. In the lab, the power draw of the fan was measured to be 160-watts, so the calculated fan energy use is:

Fan Energy Use: 1,160 kWh/year

Costs Total system costs can be divided into material costs and labor costs. A complete list of equipment is provided in Appendix . Labor costs include retrofitting the water-cooled condensers to each refrigeration unit and installing the system in the store.

For laboratory testing of the units, the retrofit of the refrigerator condensers was performed in two phases, which increased the labor costs for the retrofit. To estimate what the labor price might be we took the quoted cost to retrofit one unit, $520, and multiplied it by six units. Finally, the price of installing the system has an additional 4 hours of labor for a plumber at $95/hr, or $380. Total system costs are shown below:

Equipment Cost—$4,200

Labor Cost—$3,500

Total Cost—$7,700

UPDATED MODELFrom the laboratory analysis described in this report, the savings model can be updated as shown in

Energy Savings Cost SavingsReduced Air Conditioning Load 13,050 kWh $2,870

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Increased Heating Load -380 therms -$400Hot Water Preheat 8,250 kWh $1,820Refrigeration Efficiency Changes -3,420 kWh -$750System Energy Use -4,660 kWh -$1,030Total Energy Savings 13,220 kWh

-380 therms$2,520

Equipment Costs $4,200Installation Costs $3,500Total Installation Costs $7,700

Simple Payback 3 years

SENSITIVITY ANALYSISDue to the uncertainty inherent in translating the system from the lab to the store, a sensitivity analysis was done to determine the effect variations in the inputs of the values on the annual savings. For this analysis, the values of annual air conditioning run hours, refrigeration waste heat, coffee water use, domestic hot water use, and refrigeration energy changes were varied by +/-25%. The effect that varying each of these component has on the annual savings is shown in Figure 10.

$1,500 $1,700 $1,900 $2,100 $2,300 $2,500 $2,700 $2,900 $3,100 $3,300 $3,500

Air Conditioning Hours

Waste Heat Removed

DHW Use

Coffee Water Use

Refrigeration Energy Changes

Annual Savings

Figure 10: Sensitivity Analysis +/- 25% Current Estimates

From this analysis, it can be determined that the air conditioning hours and amount of waste heat removed have the largest impact on the annual savings. To understand how the variation of each of the components interact, a Monte Carlo simulation. For this simulation random values for each of the variables within the range described above were chosen and the resulting annual savings was calculated. The simple payback period of each of the 150 simulations are shown in the histogram in Figure 11.

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2.082.262.452.642.823.013.203.383.573.763.944.134.324.504.694.885.065.255.440

2

4

6

8

10

12

14

16

18

20

0%

10%

20%

30%

40%

50%

60%

70%

80%

90%

100%

Simple Payback (years)

Figure 11: Monte Carlo Simulation

From this simulation we can see that the simple payback period of the installation of the system is below 5 years 99% of the time and nearly 50% of the simulations resulted in a simple payback period of 3 years or less.

NEXT STEPS The purpose of the laboratory testing was to validate the model originally developed for the water loop heat recovery design, and to verify the system’s functionality. These objectives have been met. The next step in the agreement between CSU and Starbucks Coffee Company is the install the water loop heat recovery system in the 14944 Starbucks store.

The in-store implementation of the water loop heat recovery system will provide real time data for the system’s effectiveness, both in terms of HVAC savings as well as water heating savings. The in-store installation will eliminate any uncertainties associated with the differences between in-lab and in-store refrigeration performance. The final results of the in-store installations will be used to finalize a decision tool to be used to determine the system’s performance and payback period for different Starbucks stores around the country.

AREAS FOR FURTHER RESEARCHWhile the results from the lab testing show that this system is well suited for implementation in stores in southern California, the following are some ideas to help increase the efficiency of the system or make it more applicable to a wider variety of climates.

1. Add controls to direct heat back inside when needed for cooler climates.2. Investigate options to increase heat rejection in warmer climates such as evaporative pre-

cooling or cooling towers.

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3. Consider VFD controls on the pump to vary the flow rate rather than the fan during cooler weather.

4. Install time of day controls to lower loop temperature overnight to increase refrigeration efficiency when hot water not required.

APPENDIXAppendix ABuilding Energy Model To help understand the total hours per year the air conditioning system is running, a simple building energy model was created using store 14944. The first component of the store’s heat load that was considered was the contribution from the envelope. The values used for these calculations are outlined in Table 5.

Table 5: Building Energy Model Values

Roof North Wall West Wall South Wall East WallTotal Area (ft2) 2,973 700 400 700 400U (btu/h/ft2/°F) 0.61 0.72 0.7 0.7 0.7% Area Windows 0% 0% 50% 50% 0%U (btu/h/ft2/°F) - - 0.53 0.5 -Shading Coefficient - - 0.444 0.44 -

To estimate the heat gain or loss through the walls and roof the following equation was used.

Q=U × A× (T s−T i )Where

Q = Total heat gains or losses (Btu/h)U = Heat transfer rate of material (Btu/h/ft2/°F)A = Surface Area (ft2)Ts = Sol-air Temperature (see explanation below)Ti = Indoor temperature setpoint (occupied: 68°F / Unoccupied: 63°F)

The sol-air temperature is a term used to incorporate the heat gains due to solar radiation into the outdoor air temperature value. To calculate the sol-air temperature, the amount of solar radiation that is hitting the surface at a right angle must be determined. This value is a function of the time of day, the latitude of the building, and the surface orientation as shown in the following equations:

Roof :T s=To+(0.8 )×G sh

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Gsh=G× [sin δ sinφ+cos δ cosφ cosω ]

1 ASHRAE value for wood deck roofs2 ASHRAE value for stud walls3 ASHRAE value for double pane windows4 ASHRAE value for double pane windows is 0.88; reduced by 50% to account for overhang shading windows

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Wall :T s=T o+(0.8 )×Gsv

30

Gsv=G× [−sin δ cosφ cos γ+cos δ sinφcos γ cosω+cosδ sin γ sinω ]

δ=23.45×sin [360+ Nd+284365 ]

ω=(time−12)× 36024

Where

To = Outside air temperature (°F)G = Solar radiation at the earth’s surface (254 btu/h/ft2)ϕ = Latitudeγ = Orientation of wall (south = 0, east = -90, west = 90)Nd = Day number (between 1 and 365)

For example, the sol-air temperature for the south facing wall at noon on January 1 st in San Diego would be calculated as follows:

δ=23.45×sin [360+ 1+284365 ]=−23

ω=(12−12 )× 36024

=0

Gsv=254× [−sin (−23 ) cos (32.7 )cos (0 )+cos(−23)sin(32.7)cos (0)cos(0)+cos (−23)sin(0)sin(0)]=210

T s=62+(0.8 )×210

30=68

This value is then used in the heat gain calculation to determine the heat gain or loss though the wall area for this hour. Continuing the example from above:

Q=0.7× (700 ft2×50% )× (68−68 )=0 btuh

Since 50% of this wall is comprised of windows, the heat gain or loss through the windows also needs to be calculated. For windows, both the conductive heat transfer through the windows and the solar heat gain from radiation passing through the window must be taken into account. To do this the following equation is used:

Q=A [U (T o−T i )+G sv Fref SC ]Where

Fref = Baseline radiation transmission value based on single pane windows, 0.87SC = Shading coefficient to adjust for shading and different window types, 0.44

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For the south facing wall described above:

Q=700×50%× [0.7 (62−68 )+210×0.87×0.44 ]=26,600 btuh

These calculations are done for the roof and all four walls for each hour of the year the determine the heat load of the building envelope. The next component of the building’s heat load is the ventilation load. The ventilation load is calculated using the following equation:

Q=V × A×c p× ρ× (T o−T i )

Where

V = Volumetric flow rate, 0.5 CFMA = Building Area, ft2

Cp = Heat capacity of air, 0.24 Btu/lbm/°Fρ = Density of Air, 0.07 lbm/ft3

For the building in San Diego at noon on January 1st, the heat gain due to ventilation is given by:

Q=0.5×2,973×0.24×0.7× 60minhour

× (62−68 )=−8,990 btuh

Again, these calculations were done for every hour of the year to understand the contribution to the building’s heat load from ventilation.

The next component of the model is the internal heat load due to the employees and the customers. The first step was to understand the number of people in the store at various times in the day. For this estimate, we went to Google maps and pulled the graphs showing the busy times of the day. The graph for Monday’s is shown in Figure 12.

Figure 12: Number of Customers from Google Maps

By setting the maximum height from any day as 100%, each hour was assigned a percentage based on the height of the graph compared to the largest bar. Using this data an average business at each hour was determined. Based on in store observations, the number of employees at the busiest times is 8 and there are approximately 30 customers. Since the employees are standing and walking as they work, it is estimated that they produce 750 btu/h per person. The customers tend to be seated, so they are producing about 350 btu/h per person. Using the business data from google maps and the observed

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occupant number and activity level, the average heat load due to the people by hour can be determined and is shown in Figure 13.

12:00:00 AM 06:00:00 AM 12:00:00 PM 06:00:00 PM 12:00:00 AM -

2,000

4,000

6,000

8,000

10,000

12,000

14,000

Heat

Load

(Btu

/h)

Figure 13: Heat Load due to Customers and Employees

The final component of the building energy model is the load due to the equipment. The total heat produced from each piece of equipment was taken from Starbucks equipment heat survey. The equipment is outlined in Table 6.

Table 6: Equipment Heat Load

Non-Business Hours(Btu/h)

Business Hours(Btu/h)

Machine, Espresso 0 9,628Brewer - Digital dual 0 6,135

Oven Warming 0 22,726Soft Heat Server 0 691

Blender 0 3,683Scale 0 205

Grinder - Coffee 0 409Under Counter Refrigerator 942 942

Cooling Under Counter Refrigerator 471 4712 Door Under Counter Refrigerator 471 471

POS, compact 0 182Computer/Data Rack 1,364 1,364

Warewasher 0 9,760Lighting 0 8,754

Total 3,248 65,420

This table does not include the 21,777 btu/hr from the refrigeration equipment included in this experiment.

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The heat load from the envelope, ventilation, people, and equipment are added together for each hour of the year to determine whether the building is gaining heat and the air conditioning system would be running or the building is losing heat and the heating system is running. From this analysis, it is determined that a building similar to 14944 in San Diego would be in air conditioning mode 7,193 hours per year.

To test the accuracy of the model, the same calculations described above were run for the store in Fort Collins. The results of this analysis were compared to the utility bills. Figure 14 shows the estimated cooling hours per month compared to the electric bills for the store. The cooling hours peak during the same time as the electrical use peaks due to air conditioning use.

Figure 14: Fort Collins Cooling Hours vs. Electricity bills

The same procedure was done for the natural gas bill and the heating hours. These results are shown in Figure 15.

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Figure 15: Monthly Heating Hours vs. Natural Gas Bills

Again the general profile of the heating hours follows that of the natural gas use. There are many factors that can affect the energy use of a building besides its HVAC loads, so this does not definitively prove that the model is correct. It does, however, show that the model has captured the energy use trends of the building, and likely provides a reasonably accurate estimate of the number hours that the air conditioning system runs.

Appendix BHeat Rejection Calculations Each refrigerator was tested at three temperature increments for a period of 3-5 hours. During this time, two temperature loggers were placed in the system, one immediately before the unit’s condenser coil, and one immediately after. This setup provided logged data for the temperature increase in the water as a result of the heat exchange between the water loop and the unit’s refrigerant. With the temperature increase known, the heat output of the unit was calculated as follows:

Qrefrigerator=mC p∆T

Sample data, explain how measured number, variables that could effect in store (loading, efficiency due to air circulation problems)

Qrefrigerator = Heat output of refrigeration unit, Btu/hr

m = Mass flow rate of water through loop, lb/hr

C p = Specific heat capacity of water, 1.0 Btu/lb-°F

∆T = Temperature difference measured by loggers, °F

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The only variable on the right hand side of the equation that is not immediately known is the mass flow rate of the water. However, according to the pump curve, the flow rate should be between 5.43 and 5.51 gallons per minute. A value of 5.5 gallons per minute was experimentally verified in the lab. This equates to a mass flow rate of 2,752 lbs/hr of water. This value, as well as the specific heat capacity of water, are constant. Therefore, the heat output of each refrigerator is a function of the temperature difference logged in the experimental set up. The following figure displays the results of this testing on each unit:

Figure 16: Starbucks Refrigerators Heat Output

In general, the expectation would be that the heat output of each refrigerator would decrease as the entering temperature of the water increased. This is because the higher entering water temperature creates a smaller temperature differential between the water and the refrigerant, which translates to a reduction in heat transfer. With the exception of the vertical food case, all of the refrigerators abided by this expectation. The unexpected behavior of the vertical food case is likely due to the undersized condenser coil.

Appendix CFan Duty Cycle The first step in understanding the amount of time that the exterior radiator fans would need to run is measuring the radiator’s heat rejection efficiency. To do this cold water was run though the radiator with the fan on in a 70°F room and the exiting water temperature was measured. Next the fan was turned off and the resulting water temperature was logged. The results are shown in Figure 17.

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Figure 17: Heat Rejection Efficiency Test Results

From this test, we can determine the heat rejection rate of the radiator using the following equation.

UA= Q(T a−T w )

Where

Q = Heat Transfer Rate, Btu/hUA = Heat Transfer Efficiency of the Radiator, Btu/h-°FTa = Average Air Temperature over Radiator, °FTw = Average Water Temperature through Radiator, °F

As discussed in Appendix B, the temperature change over the radiator can be used to calculate the heat transfer over the radiator with the fan on and with the fan off.

Qfan on=mC p∆T=(2752 lbshr )×( 1btulb ∙℉ )× (6.5℉ )=17,916 Btuh

Qfan off=mC p∆T=(2752 lbshr )×( 1btulb∙℉ )× (0.2℉ )=623 Btuh

These two values can be input into the equation above to determine a UA value for this radiator.

UAfan on=17,916btu /h

(67.2−46.4)℉=860 btu

h ∙℉

UAfan off=623btu /h

(69.9−43.3)℉=20 btu

h ∙℉

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Using these values, the amount of heat that the radiator can reject at a given outside air temperature is calculated. Example calculations with an outside air temperature of 50°F is shown below.

Qfan on50℉=UA (T w−T a )=(860 btuh ∙℉ )× (95℉−50℉ )=38,780 Btu

h

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h

The amount of time that the fan would need to run to reject the 21,777 btu/h of waste heat produced by the refrigeration units can be found using the following equation.

D=(W−Qfan off )

(Q fanon−Q fanoff )

Continuing the example with an outside air temperature of 50°F:

D=(21,777−38,780)

(38,780−1,054 )=55%

This calculation is done for each outside air temperature. If the value is above 100% then the same calculation is done for the remaining heat to find the duty cycle of the second fan. The resulting duty cycles are then multiplied by the number of hours in the year at that outside air temperature to get the fan run time. The calculations for San Diego are shown in Table 7.

Table 7: Fan Duty Cycle Calculations

Outside Air Temperature (°F)

Hours per Year

Fan #1 Duty Cycle

Fan #2 Duty Cycle

Fan #1 Run Time (hours)

Fan #1 Run Time (hours)

50 40 55% 0% 21.97 051 97 56% 0% 54.56 052 201 58% 0% 115.81 053 215 59% 0% 126.97 054 204 61% 0% 123.55 055 252 62% 0% 156.61 056 268 64% 0% 171.01 057 298 66% 0% 195.38 058 322 67% 0% 217.06 059 360 69% 0% 249.70 060 397 71% 0% 283.55 061 375 74% 0% 276.02 062 452 76% 0% 343.16 063 624 78% 0% 489.09 064 651 81% 0% 527.30 065 440 84% 0% 368.69 0

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66 487 87% 0% 422.61 067 561 90% 0% 504.77 068 684 93% 0% 638.94 069 454 97% 0% 440.89 070 271 100% 0% 271.00 071 236 100% 3% 236.00 6.2472 231 100% 7% 231.00 16.9873 210 100% 12% 210.00 26.2274 217 100% 18% 217.00 39.2975 150 100% 24% 150.00 36.4476 62 100% 31% 62.00 19.30Total 7105 144

The resulting run time is then simply multiplied by the logged power draw of the fan, 160-watts, to get the annual energy use of the fan, 1,160 kWh per year.

Note: The heat transfer rate with the fan off is likely lower in the lab than in will be under normal operating conditions due to wind increasing air movement over the radiator. This means that the annual fan run time may be lower than estimated.

Appendix DRefrigeration Efficiency In addition to data logging the heat output of each refrigerator at different water loop temperatures, the power consumption of each unit was also logged. As mentioned previously, each refrigerator was tested at three temperature increments for a period of 3-5 hours.

Small Refrigeration Units (1 door refrigerator, 1 door freezer, 2 door refrigerator)The following figures illustrate two full refrigeration cycles for each smaller unit at each tested water loop temperature:

Figure 18: 1 Door Freezer

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Figure 19: 1 Door Refrigerator

Figure 20: 2 Door Refrigerator

In general, the refrigeration units tended to cycle more frequently when air-cooled as opposed to water-cooled. However, more frequent cycling does not necessarily indicate a greater duty cycle. At higher water loop temperatures, the cycles tended to become less frequent (except in the case of the 1 door freezer), but the loaded power duration also increased. The following figure illustrates the duty cycle of each refrigeration unit at different water loop temperatures, along with the air-cooled duty cycle:

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Figure 21: Small Refrigeration Unit Duty Cycles

Figure 21 illustrates the point at which a water-cooled condenser becomes less efficient than an air-cooled condenser for each unit. It should be noted that the air-cooled condenser data was taken at an ambient temperature of 70°F.

Large Refrigeration Units (vertical food case, horizontal food case, ice machine)The following figures illustrate a refrigeration cycle comparison between an air-cooled condenser and a water-cooled condenser for each larger refrigeration unit:

Figure 22: Horizontal Food Case

The horizontal food case performed as expected in the lab. To help interpret the figure, the average air-cooled power consumption was 0.64 kW. Again, for the air-cooled case, ambient air was at 70°F. This can be compared to the following water-cooled average power consumptions:

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Table 8: Horizontal Food Case Average Power Consumption

Water Loop Temperature, °F

Average Power Consumption, kW

76 0.6285 0.6691 0.67

The vertical food case did not perform as predicted, as seen below:

Figure 23: Vertical Food Case air-cooled case, the unit cycled as expected. However, for every water-cooled case, the unit ran constantly. As previously described, this is a likely indication that the condenser coil is undersized.

The final refrigeration unit tested in the lab was the ice machine. Due to its significantly higher heat output, a lower temperature data point for the water loop was not possible to obtain. The power consumption of the ice machine at two temperature increments is shown below:

Figure 24: Ice Machine

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From Figure 24, it can be seen that a higher temperature in the water loop results in a longer refrigeration cycle, which translates to a longer time requirement to produce the same amount of ice. However, although the cycles are longer, the average power consumption is lower at the higher temperature water loop.

According to the manufacturer, the ice machine requires 0.43 kWh per pound of ice produced, regardless of operating conditions. This value was corroborated by the experimental data at both water loop temperature settings, within 1%. This indicates that although the amount of ice that can be produced in a given day might be affected by changing the water loop temperature, the amount of energy per batch remains constant.

Refrigeration Load vs. Power Consumption A controlled experiment was set up in which a five-gallon bucket of warm water was placed in a refrigerator to act as a load. A temperature data logger was placed inside the bucket to monitor its temperature. The power consumption of the refrigeration unit was logged during the duration of the experiment. The experiment was performed over about two hours. During this time period, the refrigerator cycled six times. The following figures demonstrate the results of this experiment:

Figure 25: Refrigeration Load/Power Testing

The duty cycle of each refrigeration cycle was determined easily. The refrigerator load (figure on the right) was determined by the change in temperature of the bucket of water. From both figures, it is evident that the loaded power of the refrigerator remains constant, despite a changing load. From this experiment, it was hypothesized that the load on a refrigerator primarily affects the duty cycle, and not the loaded power (the average power would be affected similar to the duty cycle).

If true, this hypothesis was intended to be used predictively to determine a refrigerator’s operating condition. If the load on a refrigerator does not affect its loaded power, the only variable affecting loaded power would be the effectiveness of the condenser, which is controlled by the water loop temperature for a water-cooled condenser or ambient air temperature for an air-cooled condenser. However, further experimentation demonstrated that substantial refrigeration loads (leaving the doors open for extended periods, for example) can lead to increased loaded power of the unit.

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Appendix EThe following table provides a detailed list of costs for the system:

Table 9: System Component Costs

Description Number Cost Subtotal1/3 HP Coaxial Coil 1 $67.50 $681/2 HP Coaxial Coil 2 $77.50 $1551 HP Coaxial Coil 1 $137.50 $1381 1/2 HP Coaxial Coil 2 $112.50 $225Plate-and-Frame Heat Exchanger 2 $185.95 $3725 GPM Positive Displacement Pump 1 $454.25 $4541/3 HP Motor 1 $255 $25524" x 24" Water-to-Air Radiator 2 $384 $7681/4 HP Fan 2 $325 $650Installation Accessories 1 $510 $510Sensors and Controls 1 $640 $640

Total $4,234

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