One of the limitations of using refrigerant mixtures to achieve capacity modulation is that the range of capacity control and the temperature glide are both functions of the difference For applications in residential heat pumps, in boiling points of the two pure components. the temperature glide for the mixture should not exceed 30 F (16.7°C) (Radermacher 1986).
Operation at. temperature glides above 30 F (16.7°C) would result in situations where the temperature difference of the mixture exceeds that occurring in the source, resulting in a Thus, the amount of capacity control is limited by the degree to decrease in efficiency. which the boiling points of the two refrigerants can differ without causing the mixture to exceed a temperature glide of 30 F (16.7°C).
The mixture .R13Bl/R152a was selected for analysis because of its difference in boiling points for the pure components, which allowed for a moderate degree of capacity modulation without exceeding a temperature glide of 30 F (16.7°C).
In addition, sufficient data on this mixture were known to permit a determination of property data using subroutines developed by We recognize that the mixture is composed of two components Morrison and McLinden (1986).
that pose potential problems. R13B1 has been mentioned as being potentially damaging to the ozone layer, while R152a could present risks to safety due to flammability. However, since this mixture had the potential for matching all our objectives, we felt it should be tested in order to quantify the potential gains in efficiency from capacity modulation via composition shifting.
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TEST FACILITY
The test rig, depicted in Figure 3, consists of the following equipment:
Compressor
* * *
reciprocating 1.134-in3 (18.6 mL) displacement 10,000 Btu/h (2931 W) rated for R22 at -130 F (54.4°C) condensing temperature – 45 F (7.2°C) evaporating temperature – 95 F (35°C) ambient temperature
Condenser and Evaporator * * * * * * coaxial tube-in-tube, turbolators on refrigerant side 10-ft (3 m) total tube length ccunterflow arrangement 1.0-in (25 mm) O.D. refrigerant tube 0.56-in (14 mm) O.D. water tube 15,000 Btu/h (4396 W) nominal rating
Expansion Valve *
*
needle type
micrometer handle
Temperatures, pressures, and flowrates were measured on the refrigerant and water circuits. In addition, compressor wattage was measured using a watt transducer. Table 1 lists the sensors and gives their locations.
DATA ACQUISITION SYSTEM
The data acquisition system consisted of a desktop computer, a digital voltmeter, a multiprogrammer, a multiprogrammer interface, a digital clock, a printer, and a scanner. The desktop computer was configured with 62K bytes of memory and uses a high-speed programming language. Voltage data signals were read from the instrumentation by the digital voltmeter. The multiprogrammer was the master control unit for input/ output cards that performed such functions as closing relays to energize components and totalizing counts and contact closures 293
-I_________________________________________________
from instrumentation. Bidirectional communication between the multiprogrammer and the computer was performed through the multiprogrammer interface. The clock was a fully programmable system calendar and time-of-day source and also served as a stand-alone digital clock. The scanner was a channel selector that housed input connectors for analog data collection.
TEST PROCEDURES
Because the research focused on a residential heat pump application, it was desirable to test at conditions similar to those found in a conventional air-source system. However, it was not practical to test at Department of Energy rating conditions for an air-source system because the test rig was equipped with water-to-refrigerant heat exchangers. A decision was made to adjust the entering water temperatures to yield equivalent saturation pressures as those determined for an air-source system previously tested by Miller (1982).
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Testing Tools, A Report on what is Commercially Available Introduction Once an application has been developed, the developers must demonstrate that it performs the tasks for which it was designed accurately, reliably and with adequate performance. For this to be fulfilled extensive testing must be carried out and tools have been built to assist with this process. Developers have built different ...
Temperature differentials for the source and sink fluids were selected to represent a range of those typically found in present air-source designs. Table 2 provides a summary of the operating conditions for heating and cooling with equivalent air-source temperatures in brackets. Initial testing was with R22 to determine a baseline for comparison. The mixture R13B1/R152a was then tested by varying the composition in increments of 25% by weight, starting with pure R13B1 and concluding with pure R152a. Due to limitations of the test rig, testing was performed at all conditions in Table 2 for only two of the compositions, 50% R13B1/50% R152a and 25% R13B1/75% R152a. For the pure R13B1 and 75% R13B1/25% R152a, it was found that because the condensing pressures exceeded 400 psia (2756 kPa) for some of the test points, damage might occur to the test rig. For the pure R152a, water flow rates necessary to maintain the temperature differentials for some of the test points were out of the normal operating range for the flowmeters. Charging of the unit was accomplished with the use of an electronic weighing scale which automatically closed a valve to stop the flow of refrigerant once a prescribed amount was weighed into the unit. This was necessary because weighing cylinders were not available for either of the pure components. The charging procedure was to initially charge with the highboiling-point: refrigerant and then add the low-boiling-point refrigerant to yield the required composition. The charge was then checked at the 82 F (27.8°C) cooling mode test condition by adjusting the expansion valve to give minimal superheat [1 F (0.6°C)] at the evaporator exit and a low amount of subcooling [5 to 10 F (2.8 to 5.6°C)] at the condenser exit. If no subcooling could be achieved, charge was added to the system.
DATA ANALYSIS
Heating and cooling capacities were determined from measurements of the volumetric flow rate of water in the condenser and evaporator along with the temperature differential across each heat exchanger. From this information, capacities were calculated using the following equation. Q where p VA Cp AT density, lbm/ft 3 (Kg/m3 ), volumetric flow rate, gal/min (L/S), specific heat, Btu/lb m – F (KJ/Kg – C°) temperature difference, F (C°).
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Animal metabolism consists of the utilization of nutrients absorbed from the digestive tract and their catabolism as fuel for energy or their conversion into substances of the body. Metabolism is a continuous process because the molecules and even most cells of the body have brief lifetimes and are constantly replaced, while tissue as a whole maintains its characteristic structure. This constant ...
p V A Cp A T, (1)
The measured coefficient of performance (COP) could then be determined from dividing the capacity by the power input to the compressor. The pump power, which was relatively constant regardless of flow rate, was excluded in the calculation of COP to give a better indication of changes in the cycle efficiency. Compressor isentropic efficiency was calculated as the product of refrigerant mass flow rate and isentropic change in enthalpy divided by the work input to the compressor. Pressure and temperature measurements at the compressor inlet were used to determine enthalpy and entropy. The value for entropy at the compressor inlet was used to find the enthalpy at the compressor exit for an isentropic expansion. The refrigerant mass flow rate was determined by dividing the condenser capacity by the change in enthalpy across the condenser. A flowmeter 294
was not used to determine refrigerant mass flow rate because there was no space available for mounting a mass flowmeter, which would have been necessary due to the range in refrigerant density encountered in testing. Once the compressor isentropic efficiency was known, the ideal COP could then be determined. The ideal COP gives an indication of the effect of the refrigerant mixture on the cycle efficiency regardless of the changes in the efficiency of the compression process. For the purpose of explaining the differences in the ideal COP, a calculation was made of the log mean temperature difference, which is an indication of the heat exchanger performance. Finally, a comparison was made of the temperature glide of the refrigerant to the temperature glide of the source and sink to see how well the mixtures reduced the heat exchanger irreversibilities.
RESULTS
Results are presented for both heating and cooling operation in terms of the simulated airsource temperatures. This type of presentation allows an assumed building load line to be superimposed on the graphs to show what the expected savings in cycling and resistance heat could be for a system that changes the composition to follow the building load.
The Term Paper on Heat Pump
Outdoor components of a residential air-source heat pump A heat pump is a machine or device that transfers thermal energy from one location, called the “source,” which is at a lower temperature, to another location called the “sink” or “heat sink”, which is at a higher temperature. Thus, heat pumps moves thermal energy opposite to the direction that it normally ...
Heating Heating capacity versus outdoor temperature is plotted in Figure 4. The slopes of the lines for heating capacity versus outdoor temperature for pure R13B1 and R152a were estimated based on the slopes of the other lines because the limitations of the test rig made it impossible to obtain a second data point. From Figure 4, one can see that if it were possible to shift the composition to pure R13B1 at low outdoor temperatures, the capacity would be increased to 12431 Btu/h (3643 W), which is an increase of 2332 Btu/h (683 W) or 23% over that for R22 at 17 F (-8.3°C).
This would result in shifting the balance point from 30 F (16.7°C) to 25 F (3.9°C), yielding a savings in resistance heat. At higher outdoor temperatures, a shift in the composition to pure R152a would enable a lowering in the capacity to 8550 Btu/h (2506 W) at 47 F (8.3°C).
This is 33% lower than the capacity for R22 at the same temperature and would permit the unit to cycle less, thus saving energy normally lost due to start-up transients and off-cycle parasitics. Figure 4 also shows measured COP versus outdoor temperature. From the graph, one sees that as the concentration of R13B1 decreases, the COP increases. In addition, the pure refrigerants R152a and R22 have higher COPs than the mixtures, while the COP for R13B1 is lower. Because measured COP is a function of the ideal COP and the compressor isentropic efficiency, Figure 5 was plotted with all three parameters to show what was happening as the concentration of R13B1 decreases. The graph shows results for the 47 F (8.3°C) condition. Results for the 17 F (-8.3°C) condition follow similar trends. Figure 5 shows that in going from a composition of 75% R13B1 to 25% R13B1, the measured COP increases due to a 22% increase in ideal COP offsetting a 9% decrease in compressor isentropic efficiency. The pure refrigerants, R152a and R22, both had higher measured COPs than the mixtures as a result of higher values for the ideal COP. For the 17 F (-8.3°C) condition, it was determined that R13B1 had a low measured COP because of its low ideal’COP (6.69) in comparison with the other refrigerants. In an attempt to explain the trend in values for ideal COP, the log mean temperature difference (ATM), was calculated for the two-phase region in the condenser and evaporator. Also calculated was the difference in the fluid temperature differences at each end of the evaporator and condenser (ATSTREAM), which is an indication of how well the glide of the refrigerant matches the glide of the source or sink. The quantity, ATSTREAM, is defined by the follow equation. ATSTREAM – I where TR T, = inlet refrigerant temperature, F (°C), – exit water temperature, F (°C), TRO exit refrigerant temperature, F (°C), Tw – inlet water temperature, F (°C).
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Since mixtures theoretically have the ability to match the glide of the source and sink fluid, the ideal COP should be higher for the mixtures due to a decrease in the heat exchanger -T T (TRR – TW) , (2)
295
However, all the mixtures, except for the composition of 25% R13B1/75% irreversibilities. R152a, had lower ideal COPs than R22. As Table 3 shows, the ATM is higher in both the condenser and evaporator for all the mixtures as compared with that of the pure refrigerants. The increased ATM causes the pressure ratio to be increased, thus contributing to a smaller ideal COP for the mixtures. R152a had the smallest ATM, thus explaining its high ideal COP value. An analysis of the results at 17 F (-8.3°C), while not shown, reveals that R13B1 has a much higher ATM than the rest of the refrigerants, resulting in a lower ideal COP. Looking next at the ATsTREAM, Table 4 shows that it is much smaller for the mixtures, thus tending to increase the ideal COP. For the mixtures of 75% and 50% R13B1, this effect was obviously not enough to overcome the decrease in the ideal COP caused by the increased ATM. This interaction of the ATSTREAM and the ATM is illustrated in Figure 6. The 25% R13B1/75% R152a had approximately the same ideal COP as R22, thus the ATSTREAM and ATM essentially counterbalanced each other.
Cooling Cooling capacity and measured COP versus outdoor temperatures ranging from 82 F (27.8°C) to 95 F (35°C) are plotted in Figure 7 for mixtures of 50% R13B1/50% R152a and 25% R13B1/75% R152a along with the pure refrigerants R22 and R152a. The results show that the possibility exists for a reduction in the cycling losses by gradually reducing the amount of R13B1 in the mixture. At 82 F (27.8°C), assuming a shift to pure R152a could be accomplished, the capacity would decrease by 28% from 10419 Btu/h (3053 W) for R22 to 7452 Btu/h (2184 W).
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The measured COPs were all approximately the same except for R152a. Figure 8, which is similar to Figure 5, shows that the gain for R152a was mainly the result of a much higher ideal COP. A comparison of the mixtures versus R22 shows that the measured COPs were approximately equal because of only small variations in the ideal COP and compressor isentropic efficiency. An analysis of the ideal COP revealed that the mixtures were approximately the same as R22 because the increased efficiency, which was caused by the reduced heat exchanger irreversibilities, was offset by a reduction in efficiency, which was the result of an increase in the ATM. R152a had a higher ideal COP as the result of a much lower ATM than the other refrigerants.
CONCLUSIONS AND RECOMMENDATIONS
The following conclusions relate only to the particular test rig and conditions under which it was tested. The results should not be generalized to apply to all heat pumps. Different conditions, such as larger temperature changes in the heat exchangers, or different refrigerant mixtures could greatly alter the results. – The ability to change the composition of a mixture of R13Bl/R152a in order to adjust the capacity to follow the building load would lower the amount of resistance heat required at low outdoor temperatures and could reduce the cycling losses at higher outdoor temperatures. – Heat exchanger irreversibilities for the mixtures were reduced as a result of the temperature changes of the refrigerant more closely matching those of the source and sink. This was the Measured heating COPs for the mixtures were lower than those for R22. result of a larger log mean temperature difference in both heat exchangers that increased the compressor work. The larger log mean temperature difference was probably due to a degradation in the heat transfer coefficient. However, this is only speculation, since the area for two-phase heat transfer was not measured, thus allowing an accurate determination of the heat transfer coefficient. – Measured cooling COPs for the mixtures were approximately equal to those for R22 as the increase in efficiency, brought about by the reduced irreversibilities, virtually offset the decrease in efficiency, resulting from the increased log mean temperature difference. – Future research should be aimed at improving the heat transfer coefficient for mixtures, which would reduce the log mean temperature difference and thus increase the COP.
296
References
Miller, W. A. 1982. “Laboratory evaluation of the heating capacity and efficiency of a highefficiency, air-to-air heat pump with emphasis on frosting/defrosting operation”. ORNL/CON69 Oak Ridge, Tennessee: Oak Ridge National Laboratory. Morrison, G., and McLinden, M. O. 1986. “Application of a hard sphere equation of state to refrigerants and refrigerant mixtures.” NBS Technical Note 1226 Gaithersburg, Maryland: National Bureau of Standards. Radermacher, R. 1986. “Advanced versions of heat pumps with zeotropic refrigerant mixtures.” ASHRAE Transactions, Vol. 92, Part 2, pp. 52-59. Stoecker, W. F. 1978. “Improving the energy effectiveness of domestic refrigerators by the application of refrigerant mixtures.” ORNL/Sub-78/55463/1 Oak Ridge, Tennessee: Oak Ridge National Laboratory.
ACKNOWLEDGMENTS
This research was sponsored U.S. Department of Energy, Systems, Inc.
by the Office of Buildings Equipment Research and Development, under contract DE-AC05-840R21400 with Martin Marietta Energy
TABLE 1 Sensor Locations
Sensora
Circuit”
Location
T1 T2 T3 T4 T5 T6 T7 T, T9 T10 F 1 F 2 P 1 P 2 DP 1 DP 2
R R R R R R W W W W W W R R R R
Compressor discharge Condenser inlet Condenser exit Evaporator inlet Evaporator exit Compressor suction Condenser exit Condenser inlet Evaporator inlet Evaporator exit Condenser exit Evaporator exit Compressor discharge Compressor suction Condenser inlet, exit Evaporator inlet, exit
a T – Thermocouple; F – Flowmeter; P – Pressure Transducer; DP – Differential Pressure Transducer b R – Refrigerant; W – Water 297
TABLE 2 Operating Conditions for NARM Testing
Variable Evaporator inlet water temperature, F (°C) Condenser inlet water temperature, F (°C) Water A T, (evaporator),
Cooling mode 80 (26.7)
Heating mode 42 (5.6), [17 (-8.3)] 60 (15.6), [47 (8.3)] 60 (15.6)
85 (29.4), [82 (27.8)] 102 F (38.9), [95 (35)] 20 (11.1)
10 (5.6)
F (C*)
Water A T, (condenser), 20 (11.1) 20 (11.1)
F (C°)
TABLE 3 Log Mean Temperature Differences for Pure and Mixed Refrigerants at the 47 F (8.3°C) Rating Condition Refrigerant
R22
75% R13B1 25% R152a
50% R152a 50% R152a
25% R13B1 75% R152a
R152a
ATCOND
25.1 (13.9)
34.0 (18.9)
28.1 (15.6)
27.5 (15.3)
21.4 (11.9)
F (C°)
ATMAP
“EVAP
12.6
(7.0)
16.0 (8.9)
17.2 (9.6)
14.2 (7.9)
9.0 (5.0)
F (C°)
TABLE 4 Fluid Temperature Differences for the Evaporator and Condenser at the 47 F (8.3°C) Rating Condition
Refrigerant
R22
75% R13B1 25% R152a
50% R13B1 50% R152a
25% R13B1 75% R152a
R152a
AT
STREAM C16.1 ,
(9.0)
5.3 (3.0)
3.9 (2.2)
0.0 (0.0)
15.9 (8.9)
F (C°) AT
STREAM
CO
,
14.4 (8.0)
0.3
(0.2)
6.0 (3.3)
4.4 (2.4)
11.9 (6.6)
F (C)
EVAP
298
^
3
{
v
/4 Dp>~~~~ ^ ~~~~~NARM /
W
–
2′
w
–
1
2
ENTROPY (s)
Figure 1 Temperature vs. entropy diagram for pure and mixed refrigerants Figure 2
SOURCE TEMPERATURE
Capacity vs. source temperature for pure and mixed refrigerants
WATER IN
WATER OUT
WAT STORAGE TANK T WATER OUT
WATER IN
CONDENSER
REFRIGERANT OUT
COMPRESSOR EXPANSION VALVE REFRIGERANT IN
WATER OUT HEATING COIL
EVAPORATOR P
WATER IN
PROCESS WATER OUT
PU PUMP PROCESS WATER IN
Figure 3
Test rig schematic
299
OUTDOOR TEMPERATURE (°C) -17.8 20 _____20 –6.7 4.4 l________ ___ 15.6 5.9
16 –»2
–
/ y/
RESISTANCE HEAT SAVINGS
–
4.7
X
I
12 –
75% R13B1 5 R22
i
5 R13 B1 Ss CSA
r
^CYCLING VISAVINGS
Y^7 BUILDING LOAD –R152a
8
50% R13B1 25% R13B1
^
_
2.3
z
I
~~~~~~~~~~~~~_4
1.2
4.0 R152a R22 25% R13B1 50% R13B1 759b R13B1 3.0 R13B1 0
8
cc a w
3.5
~-
i
/^-
2.5
0 20 40 60 OUTDOOR TEMPERATURE (“F)
Figure 4
Heating capacity and measured COP as functions of outdoor temperature 300
4.0
4.0 C
I
I
I
I 0——————-I
ot
a 3.5 –
A
o
12.0
r-
10.0 10,0
12.o 6.0
0.45
0
_,.
CO
coujO
°c a; 0.35
0.30 75% R13B1 25% R152a 50% R13B1 50% R152a 25% R13B1 75% R152a R152a R22
REFRIGERANT COMPOSITION BY WEIGHT FRACTION
Figure 5
Heating performance parameters as functions of refrigerant composition 30 301
OUTDOOR TEMPERATURE (°C) 23.9 16 I 29.4 35.0 40.6 4.7
I
12 J~R22 Wlilt~~~~~
3.5 SAVINGS
l
21CYCLING
50% R13B1
I
cc”*””””^
R2R252a
_
8_ ^
25% R13B1
2.3 0 BUILDING LOAD
I-
4
1..
0
0
(‘4
3.0 R162a V
ENTROPY (s)
R22
g 2.5
265% R13B1 A R22 50% R13B1
MIXTURE
—–SOURCE AND SINK
1.5 75
85 95 105
///////Tm EFFICIENCY GAIN
TT EFFICIENCY
Figure 6
LOSS
Figure 7
OUTDOOR TEMPERATURE (IF)
Temperature vs. entropy diagram shoving efficiency gains and losses of mixed vs. pure refrigerant refrigerant
Cooling capacity and measured COP as functions of outdoor temperature
3.0
a-
2 .5
O 0
2.5 6.
o
2.0
8.0
I
I
I
I
aO 6.0.35
0.45
i
Q
Ia
0.40
0.35
II
0.30 50% R13B1 50% R152a
I
25% R13B1 75% R152a R152a R22
REFRIGERANT COMPOSITION BY WEIGHT FRACTION
Figure 8
Cooling performance parameters as functions of refrigerant
303