Part Load Performance of a Two-Stage Variable Capacity Air Source Heat Pump System in Cooling Mode.
In an effort to demonstrate sustainable housing technologies in Ontario, the Toronto and Region Conservation Authority (TRCA) along with the Building Industry and Land Development (BILD) Association have implemented the "Sustainable Archetype House" project at The Living City Campus at Kortright in Vaughan, Ontario, Canada. This prototype twin house is designed to demonstrate sustainable housing technologies through research, education, training, market transformation, and partnership programs. Included in the sustainable technologies is a two-stage variable capacity air source heat pump in House A. A long term monitoring system has been implemented to monitor the ASHP thermal performance using a data acquisition (DAQ) system, and analyzed using LabVIEW platform (1), (2), (3).
[FIGURE 1 OMITTED]
The use of air source heat pumps (ASHP) for efficient residential heating and cooling are greatly favoured over ground source heat pumps mainly due to the lower installed costs. However one great disadvantage of the air-source heat pump is the decrease of heat output and coefficient of performance (COP) in colder climates (4). Heating requirements in climates like Ontario and Canada will provide a challenge to the air source heat pump because of outdoor temperatures that can reach -30[degrees]C (-22[degrees]F). Also, because of such cold winter temperatures in the heating season, to meet the required building heating demand a large sized heat pump will often be used. Due to such a large capacity heat pump, the compressor will often operate at part loads to meet the building demand at milder winter temperatures and in cooling mode. This part load operation causes a reduction in efficiency and comfort due to the need of heat pump cycling. Multiple or modulating compressors address mismatched loads by sizing compressor capacity to meet heating loads at full capacity, and part load operation with the lower stage compressor to satisfy cooling loads and dehumidification. However the problem of reduced heating cycle efficiency as ambient temperature decreases still remain (5). Variable capacity ASHP's however offer potential improvements in the efficiency and reliability of operation. These improvements result from a reduction in cyclic operating time, and enhanced performance at lower capacities (6).
Numerous ASHP performance analyses have been reported, but only a few studies have looked at the part load performance. Ugursal et al. (7) has investigated the thermal performance of an air to air source heat pump installed in an R-2000 house. One year monitoring was implemented to study the performance of the house and the air source heat pump system. Data was gathered from the installed sensors every three minutes using a micro-processor based data acquisition system. The results of the study indicated that the overall seasonal cooling COP of the air source heat pump system was 1.95. In terms of part load performance, the research group noticed that the heating COP fell sharply when the outdoor temperature was above 6[degrees]C (42.8[degrees]F) because the heating requirement of the house was lower than the heating capacity of the heat pump and caused the unit to operate with short cycles in a less efficient part load mode.
De Swardt et al. (8) have experimentally monitored the performance of a single stage air to air heat pump with a nominal cooling capacity of 7.43 kW (2.1 ton) at a condensing temperature of 39.69 [degrees]C (103.4[degrees]F) and an evaporation temperature of 3.22[degrees]C (37.8[degrees]F). The results of the experiment concluded that the air to air heat pump had a cooling COP of 3.1 at an outdoor dry bulb temperature of 23[degrees]C (73.4[degrees]F).
Fadel et al. (9) investigated the performance of a variable speed drive heat pump through simulation. The simulation results demonstrated that the variable speed heat pump had improved COP at reduced frequency and higher heating rates at high frequencies. The improved COP was desirable for enhancing the part load performance of the heat pump. It was also noted that the efficiencies of the heat pump deteriorated at higher frequencies because the heat exchangers were more heavily loaded requiring the condenser to operate at a higher temperature and pressure and similarly forcing the evaporator to operate at a lower pressure.
Tassou et al. (10) have looked at the part-load performance of air-to-water heat pump systems to investigate the losses associated with compressor cycling and the use of backup heating. Experimental results have been obtained from an air-to-water heat pump designed for a maximum output of 8 kW (27. 3 mBtu/hr). The results of the study indicated that even when heat pump sizing was performed at an optimal level, the losses due to on/off cycling reduced the efficiency of the system about 6%, and backup resistance heating caused decrease in efficiency of about 4% at low ambient temperatures.
The studies that investigated the part load performance of variable capacity heat pumps in cooling or heating mode have all concluded similar findings. The above studies have all concluded that variable capacity heat pumps have improved performance at lower capacities and more reliable operation as a result of a reduction in cyclic operating time.
To analyze the thermal performance of the ASHP system using the monitoring system, data for various operating conditions was collected every 5 seconds. Data such as outdoor temperature and relative humidity, supply/return temperatures and relative humidities, supply air velocity, and the power consumption of the compressor, outdoor and indoor fan was collected. The two-stage variable capacity ASHP used for testing has a rated cooling capacity of 34,000 BTU/hr (9.96 kW) at an indoor condition of 26.7[degrees]C (80[degrees]F) dry bulb and 19.4[degrees]C (66.9[degrees]F) wet bulb temperature, and an outdoor condition of 35[degrees]C (95[degrees]F) dry bulb and 23.9 [degrees]C (75[degrees]F) wet bulb temperature. The heat pump varies its capacity by varying its compressor and outdoor fan speeds. A direct expansion coil AHU system is used for delivery of conditioned air. Table 1 lists the air temperature, relative humidity, and air velocity sensors required for ASHP analysis which provide output signals in milliamps. Table 2 lists the electricity consumption sensors of the ASHP that provide output signals in pulses.
Table 1. Air temperature, relative humidity, and air velocity sensors Sensor Name Sensors type Location RH12 Relative Humidity Main return air from zone to AHU AT12 Air Temperature Main return air from zone to AHU RH7 Relative Humidity Main supply air AHU to zone AT7 Air Temperature Main supply air AHU to zone AV1 Air Velocity Meter Supply air duct from AHU Table 2. Watt-hour sensors Sensor Name Sensors type Location 3-P-1 Watt-node Compressor & outdoor fan power 1 Watt-node AHU fan power
A schematic of the AHU is given in Figure 2 depicting the return air line, the supply air line, the connection with the ASHP and the corresponding sensors and their locations. Using the following equations, the performance of the system was analyzed
[Q.sub.Cool]=[m.sub.air]([h.sub.(a, i)] - [h.sub.(a, o)]) (1)
[COP.sub.Colling]=[Q.sub.Cool]/ [Q.sub.Electrical)] (2)
[Q.sub.Electrical)]=[W.sub.Compressor] + [ W.sub.Outdoor fan] (3)
The specific enthalpy of moist air was taken as follows:
h=[C.sub.pa]T + x([C.sub.pw]T + [h.sub.we]) (4)
[FIGURE 2 OMITTED]
RESULTS AND DISCUSION
The summer data collection was originally planned to start at the beginning of August 2010 and continue until the end of August. However, because of some issues with dysfunctional sensors, the summer data collection actually commenced on August 23 through September[14.sub.th]. During this test period, the ambient temperature range was between 15 [degrees]C (59[degrees]F) and 34 [degrees]C (93.2[degrees]F) and provided a good temperature range to properly investigate the performance of the ASHP. As mentioned earlier in the literature review, the part load characteristics of heat pumps have a great impact on the overall coefficient of performance. Larger capacity heat pumps designed to handle extreme design conditions often will have a greater frequency of compressor cycling to meet lower thermal demands. This on/off cycling causes a degradation of performance which leads to inefficient heat pump operation. Due to the fact that the heat pump will run at a lower capacity than the design condition for a significant amount of time, the effect of part load efficiency plays an important role. The variable speed compressor ASHP is designed to run for a longer period at lower speeds to meet the part loads. Figures 3 and 4 depict the operating characteristic of the two-stage variable capacity ASHP during the test period. Figure 3 illustrates the daily duration of compressor operation starting on August 23 (Day 1) through to September 14 (Day 23). Figure 4 illustrates the daily number of on/off cycling the compressor had to undergo to meet the thermal demand. From these two figures, it can be seen that the maximum cycle per day is only one, with long operating times ranging from 3 hours - 11 hours per day.
[FIGURE 3 OMITTED]
[FIGURE 4 OMITTED]
Using the data gathered from the sensors, along with the manufacturers rated conditions, the part load performance of the ASHP was investigated. Figure 5 below illustrates the part load performance of the ASHP both thorough the data collection and the manufacturers specifications, where the relationship of COP ratio and the input ratio with the heat pump capacity ratio is given. The COP ratio is defined as the instantaneous COP over the rated COP at various heat pump capacities. The input ratio is defined as the instantaneous heat pump input power over the rated input power at various heat pump capacities.
[FIGURE 5 OMITTED]
During the tested period, the ASHP capacity ratio ranged from about 52% - 57%. As a result, the experimental part load performance of the heat pump only exists in this region. The experimental COP ratio and the Input ratio obtained from the data were plotted on the manufactures part load performance curve to depict the similarities between the two. The experimental COP (COP Expr.) ratio curve in Figure 5 illustrates that at 55% of the rated capacity the heat pump COP is 25% more than the rated capacity, while the experimental input (Input Expr.) ratio curve in Figure 5 suggest that at 55% of the rated capacity the ASHP will only require 45% of the rated power. This suggest that if a single speed air source heat pump system was used instead, the compressor would often cycle on and off to meet the required load because only around 52%-57% of the full capacity was required. When comparing the experimental data points with the manufacturer's specifications, it can be seen that the experimental data falls directly on the manufacturer's part load performance curve (COP Manu., Input Manu.).
Using the results of the ASHP performance for the tested period, the typical summer seasonal performance was obtained through data extrapolation. The Toronto metropolitan temperature file in TRNSYS 16 was used to extrapolate the typical summer performance of the ASHP from May 22 - September 30. The performance extrapolation only considers the consumption of the outdoor fan and the compressor. Figure 6 depicts the average daily consumption and cooling output of the ASHP from May 22 to September 30. Figure 7 depicts the average cumulative consumption and cooling output during the same period. At the end of the cooling season the final cooling output was obtained to be 1882.8 kWh (6.42 mmBtu) and the relative electricity consumption turned out to be 428.9 kWh (1.46 mmBtu). Thus the extrapolated summer seasonal COP from these two results turned out to be 4.39.
[FIGURE 6 OMITTED]
[FIGURE 7 OMITTED]
A comprehensive monitoring system was used to investigate the cooling thermal performance of a two-stage variable capacity air source heat pump system in House A of the Archetype Sustainable Twin Houses. Data was collected from a variety of sensors for a three week period to investigate the part load cooling efficiency and compare it with the manufacturer's specification. From the collected data, it was observed that the variable speed compressor heat pump operates for longer periods at lower compressor and outdoor fan speeds to meet part loads. When only analyzing the compressor and outdoor fan, there are improvements in the efficiency and reliability of operation from a reduction in cyclic frequencies and improved performance at lower rated operating speeds/capacities. From the investigated part load performance it was noticed that at 55% of the rated capacity the heat pump COP is 25% more than the rated capacity and the ASHP will only require 45% of the rated power. When comparing with the manufacturers specifications, the experimental results from the data collection produced very similar
results. The 3 week test results were also used to extrapolate a typical summer season performance using the TRNSYS 16 Toronto metropolitan weather file. The overall seasonal cooling coefficient of performance (COP) was obtained by dividing the total cooling output by the total electricity consumption at the end of the season. The seasonal cooling COP was computed to be 4.39. Overall the findings indicate that because the heat pump system mostly operates during part loads, a variable speed compressor ASHP presents a good solution for residential cooling system due to its enhanced part load performance. If a single stage system was used for these houses, the heat pumps would be oversized causing a reduction in efficiency and comfort due to the need of heat pump cycling. Further research is currently taking place to investigate the part-and full-load performance of the variable compressor ASHP system in heating mode.
The financial support from Regional Municipality of Peel, Regional Municipality of York, City of Toronto, Building Industry and Land Development (BILD) Association, Toronto Region and Conservation Authority (TRCA), MITACS Accelerate, Reliance Home Comfort and Union Gas Ltd in implementing this project is greatly appreciated.
[m.sub.air]= Mass flow rate of air within the system (Kg/s) / (lb/s)
[h.sub.(a,o)]= Enthalpy of supply air (KJ/Kg) / (Btu/lb)
[h.sub.(a,i)]= Enthalpy of return air (KJ/Kg) / (Btu/lb)
[c.sub.pa] = Specific heat capacity of air (KJ/Kg.[degrees]C) / (Btu/lb.[degrees]R)
T = Air temperature ([degrees]C) / ([degrees]F)
[c.sub.pw] = Specific heat capacity of water vapour (KJ/Kg.[degrees] K) / (Btu/lb.[degrees] R)
[h.sub.we] = Evaporation heat of water at 0[degrees]C (KJ/Kg) / (Btu/lb)
x = Humidity ratio (Kg/Kg) / (lb/lb)
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(10). Tassou, S, C Marquand and D Wilson. "Part-Load Performance Analysis of Air-to-Water Heat Pump Systems." Journal of the Institute of Energy (1984): 364-367.
Amir A. Safa Student Member ASHRAE
Alan S. Fung Member ASHRAE, Ph.D, P.Eng
Wey H. Leong Member ASHRAE, Ph.D, P.Eng
Amir A. Safa is a MASc. student in the Department of Mechanical Engineering, Ryerson University, Toronto, Ontario. Alan S. Fung and Wey H. Leong are associate professors in the Department of Mechanical Engineering, Ryerson University, Toronto, Ontario.
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|Author:||Safa, Amir A.; Fung, Alan S.; Leong, Wey H.|
|Date:||Jul 1, 2011|
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