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    Hot water supply using a transcritical carbon dioxide heat pump : a thesis presented in fulfilment of the requirement for the degree of Master of Engineering at Massey University, Palmerston North, New Zealand
    (Massey University, 2004) Kern, Roland
    In New Zealand (NZ) a typical household uses between 160-330 I of hot water per day at 50 to 60°C. Most hot water systems are electrically heated. Heat pumps using carbon dioxide (CO₂) in the transcritical heat pump cycle offer high potential for energy savings. The use of CO₂ also offers further benefits such high volumetric heating capacity, reduced environmental impact, good availability and low costs. The objective of this project was to design, build and test a hot water supply system (HWSS) using a CO₂ heat pump. The main components of the HWSS were the heat pump, a stratified hot water storage cylinder (HWC), a water pump and a control system. The heat pump design was based on a prototype Dorin CO₂ compressor which was available. Key features were use of a vented spiral tube-in-tube heat exchanger for the gas cooler, use of a low pressure receiver incorporating an internal heat exchanger after the evaporator and the use of a back-pressure regulator as the expansion valve. The heat pump had a nominal design heating capacity of 8.1 kW with a COP of 3.9 at 0°C/34.8 bar.a evaporation temperature/pressure and 100 bar.a discharge pressure when heating water from 15°C to 60°C. The prototype heat pump performance was measured for a range of operating conditions including 0°C/33.8 bar.g to 15°C/49.8 bar.g evaporation temperatures/pressures, 18 to 30°C cold water inlet temperature, 40 to 60°C hot water outlet temperature and 90 to 120 bar.g discharge pressures. Liquid refrigerant and/or oil carry over caused by limited LPR separation capacity and/or oil foaming in the LPR was apparent for some trials but could not be completely eliminated. The compressor isentropic and volumetric efficiencies were about 30% lower than stated by the manufacturer. Possible reasons were mechanical and/or compressor oil related problems. The gas cooler was marginal in capacity especially when the heat pump operated at high evaporation pressure conditions. The measured heat pump heating capacity at the design conditions was 5.3 kW at a COP of 2.6. The heat pump COP was not sensitive to the discharge pressure across a wide range of operating conditions, so constant discharge pressure control was adopted. Overall best heat pump efficiency for 60°C hot water was achieved at 105 bar.g discharge pressure. At these discharge conditions the heating capacity and COP ranged from 4.8 kW and 2.2 at 0°C/33.8 bar.g evaporation temperature/pressure and 30°C cold water inlet temperature to 8.7 kW and 3.9 at 15°C/49.8 bar.g evaporation and 18°C water inlet respectively. A mathematical model of the HWSS was developed. The model parameters were determined from a small set of separate trials. The overall agreement between measured and the predicted HWSS performance was good. The HWSS performance was predicted for conditions likely to occur in a one or two family home. The biggest efficiency losses were HWC standing losses to the ambient air. The heat pump operated with close to the maximum COP of 2.75 because the water inlet temperature seldom rose above 25°C. There was potential for efficiency improvements if the short on/off intervals caused by the relatively small HWC relative to the heating capacity of the heat pump could be avoided. Overall, the investigation has shown that the CO₂ heat pump combined with a stratified HWC can provide a very efficient HWSS. The heat pump prototype performance was competitive with conventional heat pumps but there was significant potential for efficiency improvements due to the poor compressor performance. However, the availability and costs of heat pump components and the poor compressor performance constrain the commercial implementation.
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    Performance of a transcritical carbon dioxide heat pump for simultaneous refrigeration and water heating : a thesis presented in partial fulfilment of the requirements for the degree of Master of Technology at Massey University
    (Massey University, 1998) Yarrall, Michael George
    Many industrial processes require both refrigeration to less than 0°C and water heating to greater than 60°C. Traditional independent refrigeration and boiler systems have relatively poor energy efficiency, whilst conventional heat pumps can provide both cooling and heating but are limited in terms of the temperature lift that can be achieved. A novel heat pump using CO 2 as the refrigerant in a transcritical cycle has been proposed as a new technology that can overcome these disadvantages. The use of CO 2 as a refrigerant has many advantages. It is environmentally benign, safe, and has good thermodynamic properties, especially compared with fluorocarbons. The transcritical cycle involves evaporation of CO 2 at constant temperature and pressure below the critical point to provide refrigeration, while cooling of the CO 2 occurs at temperatures and pressures above the critical point to provide heating of water. The objective of this project was to design and construct a prototype transcritical CO 2 heat pump to simultaneously provide refrigeration and water heating, and to test its performance over a wide range of operating conditions. The prototype CO 2 heat pump had a nominal cooling capacity of 90 kW at -6°C and nominal water heating capacity of 127 kW from 10°C to 90°C. The prototype was designed to operate with a suction pressure of 30 bar and discharge pressure of 130 bar. The major components were a gas cooler, recuperator, flooded evaporator, low pressure separator/receiver, compressor, expansion valve, connecting piping and a control system. All components were standard high pressure equipment used by the natural gas processing industry. The gas cooler had a reasonably unique design to ensure close to pure counter-current heat exchange between the cooling CO 2 gas and the water being heated, both of which had relatively low flowrates. The compressor used was an open crankcase, reciprocating type with special gas seals on the piston rod to prevent CO 2 leakage. Refrigeration capacity (suction pressure) was controlled by varying the compressor speed. Water heating capacity was controlled by both using the expansion valve to control the CO 2 discharge pressure and varying the water flowrate through the gas cooler. The main problem encountered during commissioning of the prototype was CO 2 leakage through the compressor piston rod seals. Alternative sealing systems were tried, but the leakage remained an on-going problem that prevented prolonged operation of the prototype, such as would be necessary in industrial applications. Performance of the prototype was determined by energy balances based on measurements of CO 2 and water flowrate and temperature when it operated at steady-state. The energy balances generally agreed to within 6%. Trials were performed with suction pressures from 29.6 to 35.5 bar, discharge pressures from 80 to 130 bar, with hot water outlet temperatures from 65°C to 90°C, and evaporator water inlet temperatures from 11°C to 21°C. When heating water to 90°C and providing refrigeration at 1°C (35.5 bar suction pressure), the maximum overall Coefficient of Performance (COP) achieved was 5.4 at a discharge pressure of 114 bar. Below this optimum discharge pressure, the COP declined due to gas cooler heat transfer limitations (lower compressor discharge temperature led to lower temperature difference in the gas cooler and high CO 2 outlet temperature). Above the optimum, the decline in thermodynamic and compressor efficiency as pressure ratio increased caused the COP to decrease. The maximum heating and cooling capacities were about 13% less than the design values. This was attributed to the lower than expected volumetric efficiency of the compressor. The performance of the heat exchangers were generally close to the design values when allowances for lower than design water flowrates were taken into account. As expected, when suction pressure was reduced to 29.6 bar (-6°C), there was up to a 10% decrease in optimum COP as well as reduced heating and cooling capacity. When heating water to 65°C rather than 90°C, the optimum COP was about 20% higher. When suction pressure or hot water outlet temperature was decreased, the optimum discharge pressure became slightly lower due to the gas cooler heat transfer being less of a limitation on overall system performance. Addition of oil to the CO 2 did not reduce the CO 2 leakage sufficiently to allow long-term operation without recharging, and had minimum impact on the performance of the gas cooler, recuperator and compressor. However, oil fouling caused a significant drop in heat transfer performance of the evaporator. The measured prototype performance agreed well with process simulations of the equipment and with results for similar laboratory scale equipment reported in the literature. Therefore, simulations could be used to optimise component and system design with a reasonable level of confidence. It was shown that the biggest increase in COP could be achieved by improving compressor isentropic efficiency rather than increased heat exchanger size. Overall, the concept of the transcritical CO 2 heat pump for simultaneous refrigeration and water heating was proven and the required energy efficiency was sufficiently high that the heat pump is likely to be economically competitive with traditional heating and cooling systems. Further work should concentrate on improving compressor design to eliminate CO 2 leakage and to improve both isentropic and volumetric efficiency.
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    A model for improvement of water heating heat exchanger designs for residential heat pump water heaters : a thesis presented in fulfilment of the requirement for the degree of Master of Engineering at Massey University, Palmerston North, New Zealand
    (Massey University, 2010) Weerawoot, Arunwattana; Arunwattana, Weerawoot
    Heat pump water heaters are a promising technology to reduce energy use and greenhouse gas emissions. A key component is the water heating heat exchanger. Two multi-zone models of the double-wall counter-current flow heat exchanger (condenser and gas cooler models) for residential air-source heat pump water heaters were developed. These models were validated against available data in the open literature. They predicted heat exchanger size within -0.8% for a HFC-134a (with oil) condenser and within -14% for a CO2 gas cooler. The multi-zone model was significantly more accurate than one and three zone models. The models for a R410A subcritical heat pump and a CO2 transcritical heat pump were used to investigate the effect of key design parameters by varying water or refrigerant flow channel size for three water heating heat exchanger configurations: circular tube-in-tube, flat tube-on-tube, and twisted tube-in-tube. For the circular tube-in-tube configuration, refrigerant flow in the annulus (case B) performed better than refrigerant flow in the inner tube. The optimal flow channels for the circular tube-in-tube configuration case B with 0.1 mm thick air gap in the double wall were found to be di (inside diameter of the 1st tube) of 8 mm and annulus [Di (inside diameter of the 3rd tube) -d2 outside diameter of the 2nd tube)] of 1.5 mm for R410A and di of 7 mm and Di −d2 of 1.0 mm for R744. The optimal flow channels for the flat tube-on-tube configuration with b1i (major length of the refrigerant flow channel) and b2i (major length of the water flow channel) both of 9 mm were found to be a1i (minor length of the refrigerant flow channel) and a2i (minor length of the water flow channel) of 1.5 mm for R410A and a1i of 1 mm and a2i of 1.5 mm for R744. The optimal flow channels for the twisted tube-in-tube configuration were found to be di of 7.94 mm and d1 (original inside diameter of twisted tube) of 12.7 mm for R410A and di of 6.35 mm and d1 of 9.525 mm for R744. At the optimal flow channel size in each configuration, heat exchanger weight of the flat tube-on-tube was lower than the circular tube-in-tube by about 34.4% for R410A and by about 66.6% for R744. This was mainly due to elimination of the air gap resistance with the tube-on-tube configuration. Heat exchanger length, weight, and pumping power of the twisted tube-in-tube with 94% contact were significantly lower than the flat tube-on-tube by about 85%, 62%, and 97% respectively for R410A and by about 65%, 35.7%, and 98% respectively for R744. Overall, the flat tube-on-tube and the twisted tube-in-tube configurations are most promising for the water heating heat exchanger in terms of the lowest investment and running costs respectively.