The COP of the heat pump was determined by:
COP = Q^ (1)
W
comp
where Wcomp is the true power input to the compressor measured directly by the wattmeter, and Qc was determined by:
Qc = mr (h2 h) (2)
Where mr is the mass flow rate of the refrigerant, and h2 and h3 are the enthalpies at the inlet and outlet of the condenser, on the refrigerant side. Enthalpy is a function of both pressure and temperature, which are measured at each point throughout the cycle.
The natural convection flow rate, mNC can be determined by performing an energy balance across the condenser. Assuming the heat lost to the environment is negligible, as the condenser was well insulated, the energy balance is expressed as:
Qc = mr (h2 – h3) = mNCCp (T8 – T7 ) (3)
where Cp is the specific heat of the water. Rearranging Eq. 3, the natural convection flow rate is then expressed as:
An analysis was conducted on the operation of an ISAHP with a varying temperature input. The mains water temperature, and therefore initial storage tank temperature was approximately 20°C. Figure 3 shows the power delivered to the glycol through the heaters, as well as; heat transferred to the refrigerant through the evaporator; heat rejected to the natural convection loop through the condenser; and power consumed by the compressor.
Fig. 3. Variation of heat transfer rates, and compressor input power throughout the test 
As shown in Figure 3, the heat transferred from the glycol through the evaporator exceeds the heat delivered to the glycol through the auxiliary heaters. This is due to a large amount of glycol in the auxiliary heaters which are initially at a higher temperature. The heat pump unit extracts the extra energy until the heat transferred through the evaporator matches the heat input to the glycol through the heaters. Finally, the heat transfer through the evaporator is observed to lag the heater input, again due to the volume of glycol in the auxiliary heaters.
A temperature probe installed within the storage tank, with thermocouples spaced 0.15 m apart, was used to determine the level of stratification of the tank. Figure 4 shows the temperatures measured in the tank throughout the test, with T11 being the temperature measured at the top of the tank, and T20 measured at the bottom. During the test the tank was observed to stratify well, and throughout the test the temperature at the bottom of the tank, T20, remains constant at approximately 21°C. Therefore, over the course of the test, the temperature of the water entering the natural convection heat exchanger/condenser was effectively constant at 21°C.
The natural convection flow rate was also calculated over the length of the test. Figure 5 shows the natural convection flow rate of water through the condenser, plotted with the temperature of glycol delivered to the evaporator on the same graph. Consistent with the previous constant temperature tests, it was observed that higher glycol temperatures delivered to the evaporator increased the flow rate, and an increase in tank temperature decrease the flow rate. This resulted in the natural convection flow rate remaining constant for the first half of the test, and then decreasing as the glycol temperature declined, and tank temperature continued to increase.
In order to compare the dynamic operation of the heat pump with the steadystate model, the measured compressor power and heat pump COP were plotted with the simulated results on the same figure. To assist in validating the computer model, the simulation was run under the same conditions as the experiment. The model predicted the power consumed by the compressor, the evaporator and condenser heat transfer rates, and the COP of the system. Figure 6 shows the heat
transfer rates through the condenser and evaporator and Figure 7 shows the COP and compressor power curves throughout the test. Both simulated and experimental values are displayed.


COP values measured throughout the test ranged from 2.4 to 3.2, and the measured compressor power ranged from 484 to 635 W. The amount of heat transferred through the condenser to the natural convection loop ranged from 1300 to 2000 kW.
Similar to the previous tests [13] run at constant glycol temperatures, the simulation program was found to overestimate the COP and heat transferred through both heat exchangers, but predicted the compressor power consumption reasonably well. Both heat transfer rates, and the COP, were overpredicted by the simulation by approximately 12%. These results were previously attributed to an over estimation of the heat exchanger effectiveness values. Neglecting this offset, the simulation, based on a steadystate governing equations, predicts the trend of the heat pump system well.
Further testing is to be carried out, investigating wider ranges of temperatures and operating conditions. New heat exchanger relationships will be derived to better predict the effectiveness values, which will bring the model in to better agreement with the actual results. With the model refined, full year simulations in TRNSYS will be carried out to determine seasonal solar fraction values for the ISAHP system.
The experimental results matched well with the simulated results for the compressor power input, but the simulation over predicted the performance on the system. The compressor power input ranged from 484 – 635 W, and the COP of the system ranged from 2.4 – 3.2 over the duration of the test. The computer model predicted the dynamic operation of the system well, except for a 12% overestimation of performance due to the model’s effectiveness relationships. New seasonal solar fraction values and life cycle cost numbers will be calculated once the full year simulations are completed in TRNSYS.
Support for this work was provided by the Solar Buildings Research Network of Canada, the Ontario Graduate Scholarship Program, and the Natural Science and Engineering Research Council of Canada.
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