Amana Refrigeration, Inc. Brazeway, Inc. Carrier Corporation. Chrysler Corporation. Copeland Corporation. Delphi Harrison Thermal Systems. Frigidaire ...
Experimental Investigation of a Refrigerator with a Dual Temperature Evaporator H. Ge and C. W. Bullard
ACRCTR-156
September 1999
For additional information: Air Conditioning and Refrigeration Center University of Illinois Mechanical & Industrial Engineering Dept. 1206 West Green Street Urbana, IL 61801 (217) 333-3115
Prepared as part of ACRC Project 66 Refrigerator Systems Analysis C. W. Bullard, Principal Investigator
The Air Conditioning and Refrigeration Center was founded in 1988 with a grant from the estate of Richard W. Kritzer, the founder of Peerless of America Inc. A State of Illinois Technology Challenge Grant helped build the laboratory facilities. The ACRC receives continuing support from the Richard W. Kritzer Endowment and the National Science Foundation. The following organizations have also become sponsors of the Center.
Amana Refrigeration, Inc. Brazeway, Inc. Carrier Corporation Chrysler Corporation Copeland Corporation Delphi Harrison Thermal Systems Frigidaire Company General Electric Company Hill PHOENIX Honeywell, Inc. Hussmann Corporation Hydro Aluminum Adrian, Inc. Indiana Tube Corporation Lennox International, Inc. Modine Manufacturing Co. Parker Hannifin Corporation Peerless of America, Inc. The Trane Company Thermo King Corporation Visteon Automotive Systems Whirlpool Corporation York International, Inc. For additional infonnation: Air Conditioning & Refrigeration Center Mechanical & Industrial Engineering Dept. University of Illinois 1206 West Green Street Urbana IL 61801
2173333115
Table of Contents List of Tables ..................................................................................... ii List of Figures .................................................................................. .iii Abstract ............................................................................................. 1 Chapter 1. Introduction ................................................................................... l 1.1 Introduction............................................................................ l 1.2 Purpose ................................................................................. 2 1.3 Experimental refrigerator .......................................................... 2
2. System Performance Analysis ........................................................ .4 2.1 2.2 2.3 2.4
Introduction............................................................................ 4 Freezer mode .......................................................................... 4 Fresh food mode ...................................................................... 5 COP and runtime fraction .......................................................... 9
References ........................................................................................ 10 Appendix A. System Modification and Instrumentation ..................................... 11 B. Refrigerator Instrumentation ....................................................... 15 B.1 Previously done instrumentation ................................................ 15 B.2 New instrumentation .............................................................. 15
C. Cabinet Conductance Measurement .............................................. 19 C.1 C.2 C.3 C.4
System setup ........................................................................ 19 VA measurement for freezer mode ............................................. 20 VA measurement for fresh food mode ......................................... 21 VA results ............................................................................ 22
D. Capillary Tube Optimization ........................................................ 23 D.1 Optimization criterion ............................................................ 23 D.2 Freezer mode optimization ....................................................... 23 D.3 Fresh food mode optimization ................................................... 27
E. Variable Speed Evaporator Fan .................................................... 32 F. Experimental Data Sheets ............................................................. 34
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List of Tables Table
2.1 2.2 2.3 C.l D.1
Page
Freezer system performance with variable fan speeds ................................... 5 Fresh food system performance with variable fan speeds ................................ 5 COP and Runtime fraction for freezer and fresh food .................................... 9 Reverse heat leak test results ............................................................... 22 Freezer mode performance with various captube length and refrigerant charge .... 25
D.2 Captube #4 in fresh food mode ............................................................ 28 D.3 Fresh food mode optimization with a charge of 350 grams ............................ 28 F.I Freezer mode performance optimization ................................................. 35 F.2 Fresh food mode performance optimization ............................................. 38 F.3 Freezer system performance with variable fan speeds .................................. 40 F.4 Fresh food system performance with variable fan speeds .............................. 42
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List of Figures Figure
Page
2.1 2.2 2.3 2.4 2.5 2.6 Al
Subcooling and superheat vs. evaporator fan air flow rate ............................... 6 Condensing temperature vs. evaporator fan air flow rate ................................ 6 Evaporating temperature vs. evaporator fan air flow rate ................................ 7 COP vs. evaporator fan air flow rate ........................................................ 7 Cooling capacity vs. evaporator fan air flow rate ......................................... 7 FPC runtime fraction vs. evaporator fan air flow rate .................................... 8 Evaporator location diagram ............................................................... 11
A2 A3 B.1 B.2 C.1 C.2 0.1 0.2 0.3
Evaporator coil and ductwork diagram ................................................... 12 Capillary tubes configuration .............................................................. 13 Air-side evaporator instrumentation ...................................................... 16 Refrigerant-side instrumentation .......................................................... 18 Control volume for cabinet conductances during freezing cycle ...................... 20 Control volume for cabinet conductances during fresh food cycle ................... 21 Subcooling and superheat vs. charge ..................................................... 25 COP vs. charge ............................................................................... 26 Cooling capacity vs. charge ................................................................ 26 Subcooling and superheat vs. captube length ............................................ 29 COP vs. captube length ..................................................................... 29 Cooling Capacity vs. captube length ...................................................... 30 Runtime fraction vs. captube length ...................................................... 30 Evaporator fan power vs. fan input. ....................................................... 32 Evaporator air flow rate vs. fan input ..................................................... 32
D.4
0.5 0.6 0.7 E.1 E.2
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Chapter 1 Introduction Abstract A prototype refrigerator/freezer was fitted with a single evaporator, ducted so it could serve both compartments in a sequential manner. Two capillary tubes, one optimized for each compartment were attached to the suction line and controlled by a solenoid valve. A two-speed compressor operated at 3600 rpm in freezer mode and 2400 rpm when serving the food compartment. A variable speed fan was used to explore the response of the evaporator. The experimental results confirmed earlier simulation analyses, which predicted that achieving maximum energy savings (approx. 15%) would require a two-speed compressor having a high/low speed ratio of 5.
1.1 Introduction A "dual-temperature" evaporator refrigerator means a sequential cycle, single-evaporator refrigerator system. In this type of refrigerator, the same evaporator is sequentially used first to cool the freezer compartment and then the fresh food compartment.
The air flow over the
evaporator coil is directed into one compartment at a time, and is switched back and forth. The major advantage of a dual-temperature evaporator system is that while the refrigerator is running in the fresh food mode, the lower temperature lift in the evaporator can increase the steady state COP by nearly a factor of two. Theoretically, since the fresh food compartment represents about one-third of the total evaporator load and the operating efficiency in this mode can be nearly doubled, the overall system COP might be reduced by as much as 1520%. Furthermore, if the system is optimized for each compartment separately, the COP in the freezer mode can be also enhanced by using a different capillary tube or varying the compressor speed.
An additional performance improvement might result from using the relatively warm fresh food compartment air to defrost the evaporator during the off-cycle. This could greatly decrease, or possibly eliminate the auxiliary power requirements entirely, reducing the system complexity, its initial cost, and overall energy usage. In this respect the sequential cycle singleevaporator system has an advantage over dual-evaporator systems which need heaters to defrost the freezer coil.
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The disadvantage of this single-evaporator approach is the need for a solenoid valve to switch refrigerant flow between two capillary tubes; one optimized for freezer operation, the other for fresh food mode. In addition, there is a reliability risk associated with the air damper mechanism. We used a simple motor-driven prototype donated by Siebe Appliance Controls. Finally as Kelman and Bullard (1999) concluded during the design phase, a variable- or multispeed compressor with a 5: 1 turndown ratio would be needed in order to maximize energy savings obtainable from this technology.
1.2 Purpose Kelman and Bullard (1999) designed a dual-temperature evaporator refrigerator prototype on the basis of an existing 25 cubic foot, side-by-side Amana refrigerator, and analyzed the system performance by using an ACRC refrigerator simulation model, developed by Woodall and Bullard (1996, 1997) and carefully validated using data from the original system equipped with a single-speed compressor. The experimental work described in this report is performed on the prototype designed by Kelman. The capillary tube lengths and system charge were optimized experimentally for both freezer mode and fresh food mode on the criterion of highest COP, suitable condenser subcooling and evaporator superheat. The system performance of the optimized refrigerator was tested and analyzed at the standard Department of Energy test conditions. The additional experiments were also conducted over a range of evaporator fan speeds, to understand the different evaporator performances with variable-speed evaporator fan at dual evaporating temperatures.
1.3 Prototype refrigerator The experimental refrigerator was originally a 25 fe side-by-side Amana refrigerator, model SRD25S5W, with a factory suggested 5.125 ounces of R-134a. It was extensively modified to be a dual-temperature evaporator refrigerator system (details in Appendix A). The original single speed Tecumseh compressor, Model TP1390YXA, was replaced with a multi-speed Ameritech compressor, Model RV670-1. Using accelerometers mounted on the compressor shell (same method was described by Srichai and Bullard, 1997), the compressor
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speed was adjusted and operated at either 1890 rpm for fresh food mode or 3750 rpm for freezer mode, with a turndown ratio of approximately 2.0. The original evaporator was replaced by a re-designed one, and installed at a different location, in the rear of the fresh food compartment. The air flow across the evaporator can be drawn from and directed into either freezer compartment or fresh food compartment through a specially designed switched system. The original capillary tube was replaced with 4 tubes having different lengths and diameters to allow choices during optimization procedures at both modes.
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Chapter 2 System Performance Analysis 2.1 Introduction Capillary tubes were cut and charge was adjusted to optimize COP in both freezer mode and fresh food mode with the evaporator fan at its maximum speed (see Appendix D); then tests were conducted to see how the evaporator air-side heat transfer affects the system performance. Evaporator heat transfer characteristics are very important to the refrigerator's overall performance. Especially the air-side heat transfer, which consists most of the heat transfer resistance, contributes significantly to the evaporator performance. From our test in the fresh mode optimization, we found that a high superheat in the evaporator exit existed even when the subcooling satisfied the optimization criterion, because the capillary tube diameter was too small and its length was at its minimum. One possible Therefore during our tests we reduced the evaporator fan speed to increase the wetted are of the evaporator and to examine the tradeoff between fan and compressor power as they affect COP. The evaporator fan used here is a variable fan manufactured by Nidec (model TA500DC ). It is rated at 150 cfm by manufacturer while operating at its maximum speed at 48 Volts DC. Lower speeds were achieved by turning the voltage down.
2.2 Freezer mode In freezer mode, our optimization criterion is to seek both 5 OF subcooling and 5 OF superheat, so there is only moderate superheat at the evaporator exit. Tests were conducted by decreasing the input voltage by one-volt increments until the superheat disappeared. Table 1 shows the system performance as air flow rates were varied. Here, the air flow rates were calculated through dividing the cooling capacity, excluding the evaporator fan power, by the air side enthalpy change across the evaporator. The difference between the evaporator air outlet temperature and refrigerant outlet temperature is designated ~Tapproach.
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Evap. EvapFan Tr_sat: Fan Volt Power CondOut (V'J) (OF) (VDC)
47.8 47 46 45 44
10.8 10.4 9.9 9.4 9.0
Tr_sat: Airflow Sub- Super COP Cooling coaling -heat EvapOut rate Cap. (OF) (OF) (cfm) (OF) (V'J)
106.3 106.1 106.0 105.7 105.7
-16.4 -16.6 -16.6 -16.7 -15.4
48.74 47.42 46.28 44.78 43.41
13.5 13.3 13.2 13.0 6.0
10.4 10.2 9.6 8.1 2.0
1.20 1.20 1.21 1.23 1.19
217.3 217.7 218.6 217 212.8
AT approach (OF)
-0.5 -0.3 0.1 1.4 6.2
Table 2.1 Freezer system performance with variable fan speeds Table 2.1 shows that the system COP and capacity are relatively insensitive to the evaporator fan speed. Apparently the decrease in air side heat transfer coefficient offset the increase in two-phase heat transfer area, as superheat was reduced from about 10°F to 2°F.
2.3 Fresh food mode Evap EvapFan Tr_sat: Tr_sat: Fan Volt Power CondOut EvapOut (OF) (OF) (VDC) (V'J)
47.8 45 42 39 36 33 30 25 22 20
10.5 9.2 7.9 6.7 5.5 4.5 3.6 2.3 1.7 1.2
109.8 109.4 109.0 108.7 108.8 108.9 107.6 107.3 106.9 106.2
16.1 15.7 15.3 14.9 14.8 14.7 13.1 12.4 11.3 9.4
Air flow rate (cfm)
Subcoaling (OF)
Superheat (OF)
COP
50.29 48.08 45.37 42.6 39.36 36.44 33.09 27.29 22.84 20.04
5.1 4.8 4.4 4.1 3.2 2.1 2.3 1.0 0.2 0.1
27.6 27.9 27.7 26.6 22.8 18.4 20.9 13.6 6.1 2.7
1.86 1.89 1.91 1.93 1.94 1.95 1.95 1.91 1.89 1.82
Cooling Runtime fraction Cap.
(V'J) 262.9 261.7 258.4 256.4 254.8 252.8 244.4 233.9 225.7 210.8
0.056 0.055 0.055 0.055 0.056 0.057 0.056 0.059 0.062 0.066
Table 2.2 Fresh food system performance with variable fan speeds Since refrigerant charge was fixed during freezer mode optimization, superheat during fresh food mode optimization reached 27 of as we approached the 5°F subcooling constraint while optimizing COP for the fresh food mode. Therefore we could expect that variable fan speed would affect the system a lot more significant than freezer mode. Table 2 shows the system performance for fresh food mode with variable evaporator fan speed. Here the runtime fraction was calculated as the following: First the real heat load at the DoE standard condition was calculated using the VA values found for this refrigerator (see Appendix C); then this heat load was divided by the cooling capacity to get the runtime fraction. Reverse heat leak tests
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were conducted with the fan on, and indicated that the mullion VA was degraded about 50% due to imperfect sealing around the dampers. Figure 2.1 shows how subcooling and superheat varied with the evaporator air flow rate. Both subcooling and superheat decreased as the fan speed reduced. Fig. 2.1 Subcooling and Superheat vs. Fan Air Flow Rate 30.000
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48
45
42
39
36
33
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Evap Fan Air Flow Rate (cfm)
Figure 2.2 and 2.3 show the change of system pressures with evaporator fan speed. Both pressures decreased as the fan speed was reduced. Fig. 2.2 Condensing Temp. vs. Evap Fan Air Flow Rate 110.0 , . . . . . - - - - - - - - - - - - - - - - - - - - - , ~
109.5 -l------=~--------_____i-----~ -+-Tr_sat:CondOut 109.0 t---~=--------:;;:;.......---~-----~-!
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107.5 + - - - - - - - - - - - = > - = - - - - - - - - - - - i 107.0 106.5 106.0 51
48
45
42
39
36
33
30
Fan Air Flow Rate (cfm)
6
27
24
21
18
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Fig. 2.3 Evaporating Temp. vs. EvapFan Air Flow Rate
.
17.0
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16.0
~ 15.0
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~ 13.0
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~ ~
9.0 51
48
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Fan Air Flow Rate (cfm)
Fig. 2.4 COP vs. EvapFan Air Flow Rate 1.98 1.96 1.94 1.92
a.. 1.90
0
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1.88
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1.86
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1.84 1.82
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1.80 51
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Fan Air Flow Rate (cfm)
Fig. 2.5 Cooling Capacity vs. EvapFan Air Flow Rate 270
+---+...
260
~
250
Z. 240
---
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~ 230
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220
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200 51
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Fan Air Flow Rate (cfm)
7
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Fig. 2.6 FFC Runtime Fraction vs. EvapFan Air Flow Rate 0.068 0.066 t: 0
I-+- FFC Runtime Fraction I
0.064
'is
!!! 0.062
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0.060
:;::::;
0.058
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i 36
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30
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18
Fan Air Flow Rate (cfm)
Figure 2.4 - 2.6 show the COP, cooling capacity and runtime fraction change with the evaporator fan speed. The COP first increased then decreased as the fan speed reduced, and the highest COP achieved at about 35 cfm. The cooling capacity decreased as the fan speed was reduced, which caused the runtime fraction to increase. The runtime fraction for the fresh food compartment is still quite low, and could exacerbate cycling losses instead of reducing them. However if the compressor turndown ratio were 5: 1 instead of 2: 1 there would be no problem. Had the fresh food compartment's captube diameter been slightly larger (or its length shorter) there would have been no need to turn down the fan speed in fresh food mode. In our case due to the 2: 1 turndown ratio, a lower fan speed actually helped improve runtime fraction. In an optimized system, the tradeoff would be between noise and energy efficiency.
As both higher COP and higher runtime fraction are favorable to fresh food operation, the best evaporator fan speed for this prototype should produce an air flow rate of 33-36 cfm. Since fan power varies nonlinearly with the air flow rate, it is possible to achieve quieter operation in fresh food mode.
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2.4 COP and runtime fraction
Runtime fraction
COP
Freezer Mode Fresh Food Mode
simulation
test
simulation
test
1.05 1.86
1.20 1.95
0.54 0.17
0.59 0.06
Table 2.3 COP and Runtime fraction for freezer and fresh food
Table 2.3 shows the best COP and runtime fraction achieved in our test, compared with the simulation results by Kelman (1999).
The simulation model had been verified at a
compressor speed of 3600 rpm. The experimental results for freezer mode (the compressor speed was set to 3600 rpm) showed that system performance exceeded the prediction by 15 %. However, the simulation results for fresh food mode (the compressor speed was set to be 1800 rpm) the COP test results were much closer to the predicted values. The runtime data demonstrates confirms that there was substantial heat leakage through the relatively crude dampers separating the two compartments of our prototype. This led to a steady heat loss of about 30W to the freezer, compared to the 45W heat gain through the exterior walls of the fresh food compartment.
This reduced the fresh food compartment load by a factor of 3, and
shortened runtime accordingly. Mullion heat transfer is therefore one of the factors that needs to be addressed in the design of such systems. From our test results, the fresh food mode COP is over 60% higher than the freezer mode. Generally, the fresh food heat load is about one-third of the total refrigerator heat load. If this portion of heat load was removed under almost steady-state fresh food mode operation, we could get a nearly 15% higher overall average COP by using the dual temperature evaporator system, compared to the ordinary refrigerator. But this COP benefit might be greatly undermined by frequent on-off cycles in fresh food cooling for the real refrigerator, especially when we have a small fresh food runtime fraction as with our test prototype (the fresh food runtime fraction was only 5.7%). The turndown ratio of compressor used in our test was limited to a factor of 2.0. Using turndown ratios higher than 2.0 could increase both COP and runtime fraction for fresh food mode, thus result in greater energy savings. But it might cause lubrication problems, which could lead to poor reliability. However, some prototype units that are currently being tested by some
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manufacturers might have turndown ratios as high as 6.0. If this kind compressor are used in the dual temperature evaporator refrigerator, a much higher COP and better on-off cycle performance can be achieved.
References Kelman, S., "Dual Temperature Evaporator Refrigerator Design and Optimization," ACRC TR148, University of Illinois at Urbana-Champaign, 1999. Srichai, P.R. and C.W. Bullard, "Two-Speed Compressor Operation in a RefrigeratorlFreezer," ACRC TR-121, University of Illinois at Urbana-Champaign, 1997 Woodall, R.J., and C. W. Bullard, "Simulating effects of variable speed compressor on refrigerator/freezer performance," ASHRAE Transactions, 103:2,630-639, 1997.
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Appendix A System Modification and Instrumentation
:ft?
The experimental refrigerator was originally a 25
side-by-side Amana refrigerator,
model SRD25S5W, with a factory suggested 5.125 ounces ofR-134. It was heavily modified to be a dual-temperature evaporator refrigerator system.
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