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Performance of unitary heat pumps: review of ASHRAE Symposium session
Performance of unitary heat pumps: review of ASHRAE Symposium session* D. R. Tree School of Mechanical Engineering, Ray W. Herrick Laboratories, Purdue University, West Lafayette, IN 47907, USA Received 9 A u g u s t 1985
The following papers are reviewed: 1. System design optimization and validation for single-speed heat pump by S. K. Fischer and C. K. Rice, Oak Ridge National Laboratories. 2. Analysis ofon/offcyclingfor an air-to-air heat pump operating in the heating mode by W. A. Miller, Oak Ridge National Laboratories. 3. Field measured cycling, frosting and defrosting losses for a high efficiencyair source heat pump by V. D. Baxter and J. C. Moyers, Oak Ridge National Laboratories. 4. Design and available energy analysis of a heating-only residential heat pump for the Western Pacific Northwest by D. E. Elger, C. M. Reistad and S. Lang, Oregon State University. 5. A study of heat pump service life by Nance C. Lovvorn, Alabama Power Company and Carl C. Hiller, Electric Power Research. (Keywords:heat pumps;unitaryheat pumps;performance)
Performance des pompes fi chaleur individuelles: revue de la session de Symposium de I'ASHRAE Les rapports suivants sont passes en revue: 1. Optimisation de la conception des syst~mes et application fi la pompe it chaleur fi une seule vitesse ; 2. analyse du cycle par tout ou rien d'une pompe ~ chaleur air-air fonctionnant en mode de chauffaoe; 3. pertes en fonctionnement cyclique, par 9ivra9 e etd~oivrage mesur~es sur place pour une pompe ~zchaleur de grand rendement dont la source est fair; 4. conception et analyse de l'~nergie disponible d'une pompe ~ chaleur uniquement pour le chauffage de locaux r~sidentiels sur la c~te du Pacifique nord-ouest occidental; et 5. ~tude de la dur~e de vie d'une pompe d chaleur. Les noms et les addresses des auteurs se trouvent clans le sommaire anglais. (Mots cl6s: pompe fi chaleur; pompe/t chaleur individuelle; performance)
Five papers related with design, optimization and reliability of heat pumps were given during this symposium. The title of the papers and their authors are given above.
System design optimization Fischer and Rice described an optimization program they have developed. They presented some experimental verification of their computer results. Their optimization program requires two previously developed Oak Ridge National Laboratories' programs. The first is a steadystate model of the heat pump. This model uses as its main programs compressor maps, the NTU method for predicting heat exchanger performance and simple curve fits for fan performance. The author pointed out that one serious drawback to this model was the lack of refrigerant charge inventory calculation. In the conclusions the authors stated that the specific conclusion of the study was somewhat clouded by the limitations (refrigerant charge inventory)~- of the steady-state heat pump model. The second program needed was one to calculate annual performance. Very little information was given * The ASHRAE Symposium was held in Honolulu, HI, 23-26 June 1985 and was sponsored by TC7.6 Unitary Air Conditioners and Heat Pumps t" Material in parentheses added by reviewer
0140-7007/85/060367-05503.00 (*'~ 1985 Butterworth & Co (Publishers) Ltd and IIR
about this program though the authors said that additional information would soon be available. The optimization procedures used were similar to the standard method of steepest descent. Table 1, taken from their paper, is presented here to show typical results reported. The authors concluded that their optimized heat pump does not show a staggering improvement in annual performance from present day high efficiency units, but that it can be used to evaluate design trade-offs.
Analysis of on/off cycling The second paper by W. A. Miller reported the results of experimental tests conducted on a single-speed nominal 3 t split system air-to-air heat pump. The unit was operating in the heating mode. The results of a parameter study of heat pumps on-time, off-time and outdoor ambient temperature, and their effects on dynamic cycling operations are reported. One interesting point regarding the instrumentation used in these tests was the continuous measuring of refrigerant weight within the outdoor unit during both the on- and off-time period of the cycle. Cycling tests were conducted in the following modes of operation: the indoor and outdoor blower fans cycling on and off with
Rev. Int. Froid 1985 Vol 8 Novembre
367
Unitary heat pump performance: D. R. Tree Table 1 Steady-state performance of the scaled base case heat pump and the comparable optimized heat pump (reproduced from the paper by Fischer and Rice) Tableau 1 Performance en rbgime permanent de la pompe d chaleur dans les conditions de base et de la pompe d chaleur optimis~e qui lui est comparable (d'apr~s le rapport de Fischer et Rice) Ambient temperature (F)
Scaled base case
17 32 47 82 95
2.35 2.79 3.20 2.97 2.56
17 32 47 82 95
Capacity (Btu h ~) 15325 16220 21 520 21 810 28 250 28 310 30365 29 575 27610 27 260
Optimized heat pump
Difference (°/o)
COP 2.51 2.97 3.36 3.27 2.79
6.8 6.5 5.0 10.1 9.0 5.8 1.3 0.2 - 2.6 - 1.3
the compressor; continuous indoor blower fan operation; and off period isolation of the refrigerant in the indoor coil with and without continuous blower operation. A large amount of useful experimental data was presented in this paper. Figures 1 and 2 (Figures 3 and 5 in Miller's paper) are presented here as a sample of the data presented. On Figure 1 a time scale is inserted. At first glance, this may appear to be a logarithmic scale but it is the actual time after compressor shut-off. The conclusions of the paper, as stated by the author, are: 1. The control of off cycle refrigerant migration coupled with 2 min of extended indoor blower operation yields improvements in normal mode cycling COP and capacity. A seasonal analysis would be required to verify potential payback in terms of yearly energy savings. 2. Analysis of capacity loss per cycle, for tests having cycling rates of 8 min on/30 min off, showed major cycling loss due to offcycle refrigerant migration and its affect on the compressor in achieving steady-state operation. 3. The ratio of part-load COP to steady-state COP is dependent on outdoor temperature. As ambient temperature decreases, the redistribution of refrigerant to the outdoor heat exchanger and accumulator results in improvement in ratio of cycling COP to steady-state COP. 4. Continuous indoor blower operation degraded cycling performance below levels observed with normal cycling operation for load factor < ~ 5. 5. Cycling COP and capacity degradations are inversely related to heat pump percent on-time. As length of off cycle increases (decreasing percentage on-time), refrigerant dynamics increase the time required to establish condenser and evaporator steady-state operating conditions, liquid seal upstream of the outdoor throttle, and thermal mass operating equilibrium temperatures.
results from an air-to-air heat pump installed in an unoccupied single family dwelling. The heat pump was operated over a two year period. Although the house was unoccupied, a loads package simulating the pattern of energy use of a family of four was installed in the home. The house was located near Knoxville, TN, USA. Knoxville is an ideal place for using heat pumps. The area has both large heating and cooling loads, yet the winters are generally not so severe that large amounts of supplementary heat are required. In this short summary it is only possible to present a small portion of the data made available in this paper. Table 2 (Table 10 in the original paper) is a typical example. The authors pointed out that for the second year there is some difference in the cooling data but not the heating data. They attribute this difference to a small refrigerant leak. Other investigators have also found that the capacity and efficiency of the heat pump in the cooling mode are more sensitive to charge than the heating mode.
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Most of the experimental data available on the effects of cycling, frosting and defrosting on residential heat pumps were obtained in laboratory tests. The paper by Baxter and Moyers reported on the collection and analysis of
368
Int. J. Refrig. 1985 Vol 8 November
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Figure I Average heat exchanger temperature and refrigerant pressure plotted as a function of refrigerant migration to the outdoor unit, during the off cycle of an 8 min on/30 min off cycling test conducted at 50°F (10°C) outdoor temperature (reproduced from the paper by Miller) Figure 1 Temperature moyenne de l'~changeur de chaleur et pression du frigorig~ne en fonction de la migration du frigorig~ne vers l'appareil ext~rieur au cours de l'arrdt tf un essai de fonctionnement cyclique de 8 rain~30 rain d'arr~t d u n e temperature ext~rieure de IO~'C (d'apr~s /e rapport de Miller)
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2 Capacity, compressor power and refrigerant pumped to high side during the on period of an 8 min on/30 min off cycling test conducted at 50°F (10°C) outdoor temperature (reproduced from the paper by Miller) Figure 2 Puissance, puissance absorbee du compresseur et fi'igorigene refoule du c~t~ haute pression durant la p~riode de fonctionnement d'un essai de fonctionnement cyclique de 8 rain et 30 rain d'arr~t a une temperature ext~rieure de lO°C (d'appr~s le rapport de Miller) Figure
Field measured cycling losses
Time into off cycle (rain) Ii 2, 3= 4~ 5~ 2,5 ,
Unitary heat pump performance: D. R. Tree Table 2 Cooling season cyclingparameters (reproducedfrom the paper by Baxter and Moyers) Tableau 2 Paramdtres de fonctionnement cyclique en saison de refroidessement (aeaprds le rapport de Baxter et Moyers)
Month June 1982 July 1982 August 1982 1 Septemberto 14 October 1982 June 1983 July 1983 August 1983
Total time (h)"
Total operating time (h)
Percent on time, P (~)
Average Total cycling number of rate cycles (cycles h- l )
Outdoor
Indoor
Return air
714 383 381
183.4 149.2 89.7
25.7 39.0 23.5
142 137 96
0.20 0.36 0.25
77.2 (25.1) 80.1 (27.0) 77.5 (25.3)
76.8 (74.9) 78.6 (25.9) 78.8 (26.0)
75.0 (23.9) 76.3 (24.6) 76.1 (24.5)
991 641 361 502
116.7 172.7 123.8 186.4
11.8 26.9 34.3 37.1
118 108 79 99
0.12 0.17 0.22 0.20
76.1 (24.5) 80.4 (26.9) 82.6 (28.1) 83.8 (28.8)
77.5 (25.3) 77.5 (25.3) 78.4 (25.8) 77.9 (25.5)
75.6 (24.2) 74.8 (23.8) 75.7 (24.3) 75.2 (24.0)
Average temperaturesb [°F (°C)]
" Hours for which dynamic loss data are available
b During heat pump operation The authors report the following conclusions: Heating season
1. Over the eight winter months covered during the tests, defrosting was responsible for 10.2~o of the total heating season energy use (the total excluding supplemental I2r heating), frosting for 3.7~o, start-up transients for 8.5~o, and off-cycle parasitics for 3.3~o. The overall degradation in heating SPF (seasonal performance factor) was 26~o. 2. Of the defrosting loss ~ 60~o was due to the defrost tempering IEr heat used at an average rate of 29 kBtu h - 1 (8.5 kW). Evidence indicates that this rate could be reduced by ~ 10.2 kBtu h -1 (3.0 kW) without causing cooling of the conditioned space during defrosting. In addition, it was shown that ~ 27~o of all defrosts occurred, due to the time-temperature defrost controller employed, when no frost accumulation would be expected. These unnecessary defrosts accounted for at least 14~o of the time spent in defrost (and thus, 14~o of the defrost loss). Reducing the defrost tempering I2r rate and using a demand-defrost controller could have reduced the defrost penalty for the test unit by /> 30%. 3. A start-up transient period of 4 min rather than 2 min seems to more clearly approximate the time required for the test unit to reach steady-state conditions. Based on the field data for a 4 min transient period, the heating Cd (degradation coefficient) of the unit was estimated to be 0.26. 4. Rated heating SPFs for units with similar steady-state ratings to those of the field test unit averaged 1.99 compared to measured values of 1.91 and 1.98 for the 1981-82 and 1982-83 heating seasons, respectively.
Cooling season
1. The steady-state cooling performance of the test unit was degraded from the manufacturer's rated performance and from field demonstrated performance levels measured in 1980. Because of this degradation, quantification of cooling mode cycling losses is somewhat speculative. However, using the field measured normal mode performance as the steady-state base, it was determined that start-up transient losses accounted for 2.8% of the total cooling energy use and off-cycle parasitics for 4.4~. This low value for cooling-mode cyclic degradation is consistent with the very low cycling rates experienced in the
test unit (0.2-0.3 cycles h - l ) . The cooling Cd was estimated to be 0.11. 2. Rated cooling SPFs for units with steady-state performance characteristics similar to those actually demonstrated by the test unit averaged 2.34 compared to measured values of 2.41 and 2.30 for the 1982 and 1983 cooling seasons, respectively. In contrast, the rated SPF for a unit whose rated steady-state performance was similar to that of the test unit was 2.96, indicating that the steady-state performance degradation may have caused 20% degradation in the test unit's potential cooling SPF.
Design and available energy analysis
The paper by Elger et al. discussed the design of a heat pump for a climate where winter temperatures are mild and there is little need for summertime cooling. Such a climate exists in the Pacific Northwest of the USA. Under these conditions, a heating-only heat pump can be used. The paper discussed designs for optimizing the energetic performance of an air-to-air heating-only heat pump while satisfying cost constraints. The optimization was performed using a generalized reduced gradient algorithm together with the Oak Ridge National Laboratories heat pump model. This may be the same program that Fischer and Rice used in their optimization work. Another restriction placed on the optimized design was that it must be built from off-the-shelf items. All data presented for the so-called optimize design were obtained from the simulation program. No experimental data were taken. Taking into consideration the constraints imposed by designing an off-the-shelf item, the authors specified six independent factors that could be varied: evaporator frontal area, tube rows in evaporator, condenser frontal area, condenser subcooling, evaporator air flow rate and condenser air flow rate. Two constraints were also enforced: a total heat exchange area which would give desired capacity, and fan speed to be within the range of off-the-shelf items. Table 3 (Table i in original paper) gives the specifications of the heating-only split-type heat pump designed. This heat pump differs from a conversion heat pump in the following ways: 1. the compressor is located in the indoor unit; 2. there is no reversing valve; 3. there is only one expansion device; 4. there are smaller fan units for
Rev. Int. Froid 1985 Vol 8 Novembre
369
Unitary heat pump performance." D. R. Tree Table 3 Specifications of the heating-only beat pump final design (reproduced from the paper by Elger et al.) Tableau 3 Specifications pour la conception finale de la pompe d chaleur pour le chauffage seul (craprOs le rapport d'Elger et al.) Compressor unit Type: hermetic, reciprocating, high efficiency heat pump compressor Refrigerant: R22 Displacement: 3.516 in 3 rev - 1 (57.62 cm 3 rev- 1) Motor Speed: 3450 rev min Condenser heat exchanger Type: Finned tube with continuous, wavy, aluminium plate fins: staggered copper refrigerant tubing Frontal area: 4.774 ft 2 (0.44 m 2) Number of tube rows: 4 Number of parallel refrigerant circuits: 3 Fins: pitch 16 fins in- 1 (6.3 fins cm-1); thickness 0.006 in (0.01524 cm) Tube: nominal o.d. 5/16 in (0.794 cm); wall thickness 0.016 in (0.0406 cm); row spacing 0.625 in (1.588 cm); vertical tube spacing 1.0 in (2.54 cm) Subcooling: 17.1F at 40.7'F (9.5°C at 4.8~'C) ambient air temperature Evaporator heat exchanger Type: Finned tube with continuous, wavy, aluminium plate fins: staggered copper refrigerant tubing Frontal area: 9.99 ft 2 (0.927 m 2) Number of tube rows: 2 Number of parallel refrigerant circuits: 7 Fins: pitch 8 fins in-1 (3.15 fins cm-~); thickness 0.006 in (0.01524 cm) Tube: nominal o.d. 5/16 in (0.794 cm); wall thickness 0.016 in (0.0406 cm); row spacing 0.625 in (1.588 cm): vertical tube spacing 1 in (2.54 cm) Condenser fan unit Type: Direct drive centrifugal blower driven by an eight-pole high efficiency PSC motor Volumetric air flow rate: 1180 ft 3 min- ~ (0.528 m 3 s - 1) Static pressure drop: 0.419 in H 2 0 (104.4 Pa) Motor: speed 825 rev min ~; rated power 1/5 hp (149 W) Fan: 9.5 × 9.5 in (0.241 × 0.241 m) blower Evaporator fan unit Type: Direct drive propeller fan driven by a six-pole high efficiency PSC motor, the motor should be totally enclosed by a case Volumetric air flow rate: 2310ft 3 min t (1.09 m 3 s -1) Static pressure drop: 0.064 in H 2 0 (15.94 Pa) Motor: speed 1000 rev min: rated power: 1/5 hp (149 W) Fan : 4 blade, 1.67 ft (0.508 m) propeller fan Thermal expansion valve Nominal capacity: 2 t (7.03 kW) Rated operating superheat: 11'F (6.11'C) Static superheat: 6~'F (3.33 C) Permanent bleed factor: 1.15
both the indoor and outdoor fans; 5. there is greater fin spacing and fewer tube rows in the evaporator; and 6. a significantly greater fraction of the total air side-heat transfer area is located on the condenser side. In general, this heating-only heat pump is much more efficient (COPH (coefficient of performance for heating mode of a heat pump)=3.8 at 47F) than the average commercial unit (COPH ---2.7 at 47F), and slightly more efficient than some of the high efficient commercial available heat pumps (COPH = 3.1 at 47F). Table 4 (Table 2 in original paper) summarizes how these improvements arise. In addition to given design details, the paper presented yearly hour-by-hour availability data. Heat pump service life
There has been and still is a need for reliable data on the reliability and useful life of heat pumps. The paper by Lovvorn and Hiller provided some useful information about replacement times of heat pumps in Alabama, USA. The weather in Alabama is such that it produces very large summer loads and large winter loads. This paper reported on the replacement time of a selected set of heat pumps. The authors identified the replacement time with the heat pump service live. The heat pumps selected in the study were installed between 1964 and 1974 and were covered by a heat pump
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Int. J. Refrig. 1985 Vol 8 November
maintenance agreement for the first 10 years of their life. The maintenance contract insured that the units were properly installed and that maintenance was performed as Table 4 Summary of performance improvements for the heating-only heat pump (reproduced from the paper by Elger et al.) Tableau 4 Rdcapitulation des ameliorations de la perjbrmance de.s pompes d chaleur pour le chauffage seul (d'aprds le rapport d'EIger et al.)
Design of unit Recovery of discharge line heat losses Recovery of compressor shell heat losses No reversing valve Improved efficiency compressor Improved efficiency fan units Optimum component matching for heating: Condenser sizing Fan sizing Fewer evaporator-coil tube rows Sizing relative to space load Lower balance point temperature appropriate to heating-only unit
Increase in COPH at 47F (8.3 'C) (!~,)" 4.5 2 9 1.4 I0 5 15 5 10 ~-20"~,
SPFH increased 3 7!!11
" The percentage increase in C O P H values are not intended to be summed and multiplied by the COPH value for a commercial heat pump. Rather, they are estimates of the merits of each of the design aspects considered. Obviously, the better performance commerical heat pumps will have already incorporated some of these advantages to a certain extent
Unitary heat pump performance: D. R. Tree collected by telephone interviews from people owning the heat pumps. It should be stressed that the authors definitions of service life does not mean that there was no component failure during this period. The authors stated that additional information was available on maintenance history of the heat pumps. The conclusions of the paper were:
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14 15 16 Age (years)
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Figure 3 Heat pumps remaining in service plotted against age in years. Age 10= 10.0-10.9, age l 1 = 11.0-11.9, etc. (reproduced from the paper by Lovvorn and Hiller) Figure 3 Pompes dt chaleur encore en service en fonction de rdge (en ann~es) (d'aprds le rapport de Lovvorn et Hiller)
needed for the first 10 years. Figure 3 (Figure 1 in original paper) shows the percentage of units remaining in service as a function of years since installation. Data were
1. Of the respondents surveyed 96.4% were identified as still having heat pumps. 2. A large percentage of the original heat pump sample is still in operation, with > 50% of the 20 year old units 75% of the 15 year old units, and nearly 100°/0 of 10 year old units still in active use. 3. The median age for replacement (age at which 50°/0 of the units have been removed from service) in Alabama was -~ 20 years. 4. There was no convincing differences in service life between younger and older units, mainly due to the types of factors that affected the replacement decision. 5. More than 50% of units that were replaced were still fully operational at the time of replacement. Such replacements appear to have been motivated both by the perception of expected life, and by marketing and promotional efforts of dealer/contractors and the local utility.
Rev. Int. Froid 1985 Vol 8 Novembre
371
The following papers are reviewed: 1. System design optimization and validation for single-speed heat pump by S. K. Fischer and C. K. Rice, Oak Ridge National Laboratories. 2. Analysis ofon/offcyclingfor an air-to-air heat pump operating in the heating mode by W. A. Miller, Oak Ridge National Laboratories. 3. Field measured cycling, frosting and defrosting losses for a high efficiencyair source heat pump by V. D. Baxter and J. C. Moyers, Oak Ridge National Laboratories. 4. Design and available energy analysis of a heating-only residential heat pump for the Western Pacific Northwest by D. E. Elger, C. M. Reistad and S. Lang, Oregon State University. 5. A study of heat pump service life by Nance C. Lovvorn, Alabama Power Company and Carl C. Hiller, Electric Power Research. (Keywords:heat pumps;unitaryheat pumps;performance)
Performance des pompes fi chaleur individuelles: revue de la session de Symposium de I'ASHRAE Les rapports suivants sont passes en revue: 1. Optimisation de la conception des syst~mes et application fi la pompe it chaleur fi une seule vitesse ; 2. analyse du cycle par tout ou rien d'une pompe ~ chaleur air-air fonctionnant en mode de chauffaoe; 3. pertes en fonctionnement cyclique, par 9ivra9 e etd~oivrage mesur~es sur place pour une pompe ~zchaleur de grand rendement dont la source est fair; 4. conception et analyse de l'~nergie disponible d'une pompe ~ chaleur uniquement pour le chauffage de locaux r~sidentiels sur la c~te du Pacifique nord-ouest occidental; et 5. ~tude de la dur~e de vie d'une pompe d chaleur. Les noms et les addresses des auteurs se trouvent clans le sommaire anglais. (Mots cl6s: pompe fi chaleur; pompe/t chaleur individuelle; performance)
Five papers related with design, optimization and reliability of heat pumps were given during this symposium. The title of the papers and their authors are given above.
System design optimization Fischer and Rice described an optimization program they have developed. They presented some experimental verification of their computer results. Their optimization program requires two previously developed Oak Ridge National Laboratories' programs. The first is a steadystate model of the heat pump. This model uses as its main programs compressor maps, the NTU method for predicting heat exchanger performance and simple curve fits for fan performance. The author pointed out that one serious drawback to this model was the lack of refrigerant charge inventory calculation. In the conclusions the authors stated that the specific conclusion of the study was somewhat clouded by the limitations (refrigerant charge inventory)~- of the steady-state heat pump model. The second program needed was one to calculate annual performance. Very little information was given * The ASHRAE Symposium was held in Honolulu, HI, 23-26 June 1985 and was sponsored by TC7.6 Unitary Air Conditioners and Heat Pumps t" Material in parentheses added by reviewer
0140-7007/85/060367-05503.00 (*'~ 1985 Butterworth & Co (Publishers) Ltd and IIR
about this program though the authors said that additional information would soon be available. The optimization procedures used were similar to the standard method of steepest descent. Table 1, taken from their paper, is presented here to show typical results reported. The authors concluded that their optimized heat pump does not show a staggering improvement in annual performance from present day high efficiency units, but that it can be used to evaluate design trade-offs.
Analysis of on/off cycling The second paper by W. A. Miller reported the results of experimental tests conducted on a single-speed nominal 3 t split system air-to-air heat pump. The unit was operating in the heating mode. The results of a parameter study of heat pumps on-time, off-time and outdoor ambient temperature, and their effects on dynamic cycling operations are reported. One interesting point regarding the instrumentation used in these tests was the continuous measuring of refrigerant weight within the outdoor unit during both the on- and off-time period of the cycle. Cycling tests were conducted in the following modes of operation: the indoor and outdoor blower fans cycling on and off with
Rev. Int. Froid 1985 Vol 8 Novembre
367
Unitary heat pump performance: D. R. Tree Table 1 Steady-state performance of the scaled base case heat pump and the comparable optimized heat pump (reproduced from the paper by Fischer and Rice) Tableau 1 Performance en rbgime permanent de la pompe d chaleur dans les conditions de base et de la pompe d chaleur optimis~e qui lui est comparable (d'apr~s le rapport de Fischer et Rice) Ambient temperature (F)
Scaled base case
17 32 47 82 95
2.35 2.79 3.20 2.97 2.56
17 32 47 82 95
Capacity (Btu h ~) 15325 16220 21 520 21 810 28 250 28 310 30365 29 575 27610 27 260
Optimized heat pump
Difference (°/o)
COP 2.51 2.97 3.36 3.27 2.79
6.8 6.5 5.0 10.1 9.0 5.8 1.3 0.2 - 2.6 - 1.3
the compressor; continuous indoor blower fan operation; and off period isolation of the refrigerant in the indoor coil with and without continuous blower operation. A large amount of useful experimental data was presented in this paper. Figures 1 and 2 (Figures 3 and 5 in Miller's paper) are presented here as a sample of the data presented. On Figure 1 a time scale is inserted. At first glance, this may appear to be a logarithmic scale but it is the actual time after compressor shut-off. The conclusions of the paper, as stated by the author, are: 1. The control of off cycle refrigerant migration coupled with 2 min of extended indoor blower operation yields improvements in normal mode cycling COP and capacity. A seasonal analysis would be required to verify potential payback in terms of yearly energy savings. 2. Analysis of capacity loss per cycle, for tests having cycling rates of 8 min on/30 min off, showed major cycling loss due to offcycle refrigerant migration and its affect on the compressor in achieving steady-state operation. 3. The ratio of part-load COP to steady-state COP is dependent on outdoor temperature. As ambient temperature decreases, the redistribution of refrigerant to the outdoor heat exchanger and accumulator results in improvement in ratio of cycling COP to steady-state COP. 4. Continuous indoor blower operation degraded cycling performance below levels observed with normal cycling operation for load factor < ~ 5. 5. Cycling COP and capacity degradations are inversely related to heat pump percent on-time. As length of off cycle increases (decreasing percentage on-time), refrigerant dynamics increase the time required to establish condenser and evaporator steady-state operating conditions, liquid seal upstream of the outdoor throttle, and thermal mass operating equilibrium temperatures.
results from an air-to-air heat pump installed in an unoccupied single family dwelling. The heat pump was operated over a two year period. Although the house was unoccupied, a loads package simulating the pattern of energy use of a family of four was installed in the home. The house was located near Knoxville, TN, USA. Knoxville is an ideal place for using heat pumps. The area has both large heating and cooling loads, yet the winters are generally not so severe that large amounts of supplementary heat are required. In this short summary it is only possible to present a small portion of the data made available in this paper. Table 2 (Table 10 in the original paper) is a typical example. The authors pointed out that for the second year there is some difference in the cooling data but not the heating data. They attribute this difference to a small refrigerant leak. Other investigators have also found that the capacity and efficiency of the heat pump in the cooling mode are more sensitive to charge than the heating mode.
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Most of the experimental data available on the effects of cycling, frosting and defrosting on residential heat pumps were obtained in laboratory tests. The paper by Baxter and Moyers reported on the collection and analysis of
368
Int. J. Refrig. 1985 Vol 8 November
4.4
551.6
oL
413.7 - '~
I 2.0
0
I I 4.0 6D R-22 migration (Ib)
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Figure I Average heat exchanger temperature and refrigerant pressure plotted as a function of refrigerant migration to the outdoor unit, during the off cycle of an 8 min on/30 min off cycling test conducted at 50°F (10°C) outdoor temperature (reproduced from the paper by Miller) Figure 1 Temperature moyenne de l'~changeur de chaleur et pression du frigorig~ne en fonction de la migration du frigorig~ne vers l'appareil ext~rieur au cours de l'arrdt tf un essai de fonctionnement cyclique de 8 rain~30 rain d'arr~t d u n e temperature ext~rieure de IO~'C (d'apr~s /e rapport de Miller)
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2 Capacity, compressor power and refrigerant pumped to high side during the on period of an 8 min on/30 min off cycling test conducted at 50°F (10°C) outdoor temperature (reproduced from the paper by Miller) Figure 2 Puissance, puissance absorbee du compresseur et fi'igorigene refoule du c~t~ haute pression durant la p~riode de fonctionnement d'un essai de fonctionnement cyclique de 8 rain et 30 rain d'arr~t a une temperature ext~rieure de lO°C (d'appr~s le rapport de Miller) Figure
Field measured cycling losses
Time into off cycle (rain) Ii 2, 3= 4~ 5~ 2,5 ,
Unitary heat pump performance: D. R. Tree Table 2 Cooling season cyclingparameters (reproducedfrom the paper by Baxter and Moyers) Tableau 2 Paramdtres de fonctionnement cyclique en saison de refroidessement (aeaprds le rapport de Baxter et Moyers)
Month June 1982 July 1982 August 1982 1 Septemberto 14 October 1982 June 1983 July 1983 August 1983
Total time (h)"
Total operating time (h)
Percent on time, P (~)
Average Total cycling number of rate cycles (cycles h- l )
Outdoor
Indoor
Return air
714 383 381
183.4 149.2 89.7
25.7 39.0 23.5
142 137 96
0.20 0.36 0.25
77.2 (25.1) 80.1 (27.0) 77.5 (25.3)
76.8 (74.9) 78.6 (25.9) 78.8 (26.0)
75.0 (23.9) 76.3 (24.6) 76.1 (24.5)
991 641 361 502
116.7 172.7 123.8 186.4
11.8 26.9 34.3 37.1
118 108 79 99
0.12 0.17 0.22 0.20
76.1 (24.5) 80.4 (26.9) 82.6 (28.1) 83.8 (28.8)
77.5 (25.3) 77.5 (25.3) 78.4 (25.8) 77.9 (25.5)
75.6 (24.2) 74.8 (23.8) 75.7 (24.3) 75.2 (24.0)
Average temperaturesb [°F (°C)]
" Hours for which dynamic loss data are available
b During heat pump operation The authors report the following conclusions: Heating season
1. Over the eight winter months covered during the tests, defrosting was responsible for 10.2~o of the total heating season energy use (the total excluding supplemental I2r heating), frosting for 3.7~o, start-up transients for 8.5~o, and off-cycle parasitics for 3.3~o. The overall degradation in heating SPF (seasonal performance factor) was 26~o. 2. Of the defrosting loss ~ 60~o was due to the defrost tempering IEr heat used at an average rate of 29 kBtu h - 1 (8.5 kW). Evidence indicates that this rate could be reduced by ~ 10.2 kBtu h -1 (3.0 kW) without causing cooling of the conditioned space during defrosting. In addition, it was shown that ~ 27~o of all defrosts occurred, due to the time-temperature defrost controller employed, when no frost accumulation would be expected. These unnecessary defrosts accounted for at least 14~o of the time spent in defrost (and thus, 14~o of the defrost loss). Reducing the defrost tempering I2r rate and using a demand-defrost controller could have reduced the defrost penalty for the test unit by /> 30%. 3. A start-up transient period of 4 min rather than 2 min seems to more clearly approximate the time required for the test unit to reach steady-state conditions. Based on the field data for a 4 min transient period, the heating Cd (degradation coefficient) of the unit was estimated to be 0.26. 4. Rated heating SPFs for units with similar steady-state ratings to those of the field test unit averaged 1.99 compared to measured values of 1.91 and 1.98 for the 1981-82 and 1982-83 heating seasons, respectively.
Cooling season
1. The steady-state cooling performance of the test unit was degraded from the manufacturer's rated performance and from field demonstrated performance levels measured in 1980. Because of this degradation, quantification of cooling mode cycling losses is somewhat speculative. However, using the field measured normal mode performance as the steady-state base, it was determined that start-up transient losses accounted for 2.8% of the total cooling energy use and off-cycle parasitics for 4.4~. This low value for cooling-mode cyclic degradation is consistent with the very low cycling rates experienced in the
test unit (0.2-0.3 cycles h - l ) . The cooling Cd was estimated to be 0.11. 2. Rated cooling SPFs for units with steady-state performance characteristics similar to those actually demonstrated by the test unit averaged 2.34 compared to measured values of 2.41 and 2.30 for the 1982 and 1983 cooling seasons, respectively. In contrast, the rated SPF for a unit whose rated steady-state performance was similar to that of the test unit was 2.96, indicating that the steady-state performance degradation may have caused 20% degradation in the test unit's potential cooling SPF.
Design and available energy analysis
The paper by Elger et al. discussed the design of a heat pump for a climate where winter temperatures are mild and there is little need for summertime cooling. Such a climate exists in the Pacific Northwest of the USA. Under these conditions, a heating-only heat pump can be used. The paper discussed designs for optimizing the energetic performance of an air-to-air heating-only heat pump while satisfying cost constraints. The optimization was performed using a generalized reduced gradient algorithm together with the Oak Ridge National Laboratories heat pump model. This may be the same program that Fischer and Rice used in their optimization work. Another restriction placed on the optimized design was that it must be built from off-the-shelf items. All data presented for the so-called optimize design were obtained from the simulation program. No experimental data were taken. Taking into consideration the constraints imposed by designing an off-the-shelf item, the authors specified six independent factors that could be varied: evaporator frontal area, tube rows in evaporator, condenser frontal area, condenser subcooling, evaporator air flow rate and condenser air flow rate. Two constraints were also enforced: a total heat exchange area which would give desired capacity, and fan speed to be within the range of off-the-shelf items. Table 3 (Table i in original paper) gives the specifications of the heating-only split-type heat pump designed. This heat pump differs from a conversion heat pump in the following ways: 1. the compressor is located in the indoor unit; 2. there is no reversing valve; 3. there is only one expansion device; 4. there are smaller fan units for
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Unitary heat pump performance." D. R. Tree Table 3 Specifications of the heating-only beat pump final design (reproduced from the paper by Elger et al.) Tableau 3 Specifications pour la conception finale de la pompe d chaleur pour le chauffage seul (craprOs le rapport d'Elger et al.) Compressor unit Type: hermetic, reciprocating, high efficiency heat pump compressor Refrigerant: R22 Displacement: 3.516 in 3 rev - 1 (57.62 cm 3 rev- 1) Motor Speed: 3450 rev min Condenser heat exchanger Type: Finned tube with continuous, wavy, aluminium plate fins: staggered copper refrigerant tubing Frontal area: 4.774 ft 2 (0.44 m 2) Number of tube rows: 4 Number of parallel refrigerant circuits: 3 Fins: pitch 16 fins in- 1 (6.3 fins cm-1); thickness 0.006 in (0.01524 cm) Tube: nominal o.d. 5/16 in (0.794 cm); wall thickness 0.016 in (0.0406 cm); row spacing 0.625 in (1.588 cm); vertical tube spacing 1.0 in (2.54 cm) Subcooling: 17.1F at 40.7'F (9.5°C at 4.8~'C) ambient air temperature Evaporator heat exchanger Type: Finned tube with continuous, wavy, aluminium plate fins: staggered copper refrigerant tubing Frontal area: 9.99 ft 2 (0.927 m 2) Number of tube rows: 2 Number of parallel refrigerant circuits: 7 Fins: pitch 8 fins in-1 (3.15 fins cm-~); thickness 0.006 in (0.01524 cm) Tube: nominal o.d. 5/16 in (0.794 cm); wall thickness 0.016 in (0.0406 cm); row spacing 0.625 in (1.588 cm): vertical tube spacing 1 in (2.54 cm) Condenser fan unit Type: Direct drive centrifugal blower driven by an eight-pole high efficiency PSC motor Volumetric air flow rate: 1180 ft 3 min- ~ (0.528 m 3 s - 1) Static pressure drop: 0.419 in H 2 0 (104.4 Pa) Motor: speed 825 rev min ~; rated power 1/5 hp (149 W) Fan: 9.5 × 9.5 in (0.241 × 0.241 m) blower Evaporator fan unit Type: Direct drive propeller fan driven by a six-pole high efficiency PSC motor, the motor should be totally enclosed by a case Volumetric air flow rate: 2310ft 3 min t (1.09 m 3 s -1) Static pressure drop: 0.064 in H 2 0 (15.94 Pa) Motor: speed 1000 rev min: rated power: 1/5 hp (149 W) Fan : 4 blade, 1.67 ft (0.508 m) propeller fan Thermal expansion valve Nominal capacity: 2 t (7.03 kW) Rated operating superheat: 11'F (6.11'C) Static superheat: 6~'F (3.33 C) Permanent bleed factor: 1.15
both the indoor and outdoor fans; 5. there is greater fin spacing and fewer tube rows in the evaporator; and 6. a significantly greater fraction of the total air side-heat transfer area is located on the condenser side. In general, this heating-only heat pump is much more efficient (COPH (coefficient of performance for heating mode of a heat pump)=3.8 at 47F) than the average commercial unit (COPH ---2.7 at 47F), and slightly more efficient than some of the high efficient commercial available heat pumps (COPH = 3.1 at 47F). Table 4 (Table 2 in original paper) summarizes how these improvements arise. In addition to given design details, the paper presented yearly hour-by-hour availability data. Heat pump service life
There has been and still is a need for reliable data on the reliability and useful life of heat pumps. The paper by Lovvorn and Hiller provided some useful information about replacement times of heat pumps in Alabama, USA. The weather in Alabama is such that it produces very large summer loads and large winter loads. This paper reported on the replacement time of a selected set of heat pumps. The authors identified the replacement time with the heat pump service live. The heat pumps selected in the study were installed between 1964 and 1974 and were covered by a heat pump
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Int. J. Refrig. 1985 Vol 8 November
maintenance agreement for the first 10 years of their life. The maintenance contract insured that the units were properly installed and that maintenance was performed as Table 4 Summary of performance improvements for the heating-only heat pump (reproduced from the paper by Elger et al.) Tableau 4 Rdcapitulation des ameliorations de la perjbrmance de.s pompes d chaleur pour le chauffage seul (d'aprds le rapport d'EIger et al.)
Design of unit Recovery of discharge line heat losses Recovery of compressor shell heat losses No reversing valve Improved efficiency compressor Improved efficiency fan units Optimum component matching for heating: Condenser sizing Fan sizing Fewer evaporator-coil tube rows Sizing relative to space load Lower balance point temperature appropriate to heating-only unit
Increase in COPH at 47F (8.3 'C) (!~,)" 4.5 2 9 1.4 I0 5 15 5 10 ~-20"~,
SPFH increased 3 7!!11
" The percentage increase in C O P H values are not intended to be summed and multiplied by the COPH value for a commercial heat pump. Rather, they are estimates of the merits of each of the design aspects considered. Obviously, the better performance commerical heat pumps will have already incorporated some of these advantages to a certain extent
Unitary heat pump performance: D. R. Tree collected by telephone interviews from people owning the heat pumps. It should be stressed that the authors definitions of service life does not mean that there was no component failure during this period. The authors stated that additional information was available on maintenance history of the heat pumps. The conclusions of the paper were:
I00~
9o "~ 8o "~ 70 " ; 60
50 E
E Q.
40 3O
"6 2 0 I00
I
IO
I
II
I
12
I
13
I
I
I
14 15 16 Age (years)
I
17
I
18
I
19
20
Figure 3 Heat pumps remaining in service plotted against age in years. Age 10= 10.0-10.9, age l 1 = 11.0-11.9, etc. (reproduced from the paper by Lovvorn and Hiller) Figure 3 Pompes dt chaleur encore en service en fonction de rdge (en ann~es) (d'aprds le rapport de Lovvorn et Hiller)
needed for the first 10 years. Figure 3 (Figure 1 in original paper) shows the percentage of units remaining in service as a function of years since installation. Data were
1. Of the respondents surveyed 96.4% were identified as still having heat pumps. 2. A large percentage of the original heat pump sample is still in operation, with > 50% of the 20 year old units 75% of the 15 year old units, and nearly 100°/0 of 10 year old units still in active use. 3. The median age for replacement (age at which 50°/0 of the units have been removed from service) in Alabama was -~ 20 years. 4. There was no convincing differences in service life between younger and older units, mainly due to the types of factors that affected the replacement decision. 5. More than 50% of units that were replaced were still fully operational at the time of replacement. Such replacements appear to have been motivated both by the perception of expected life, and by marketing and promotional efforts of dealer/contractors and the local utility.
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