Design and simulation of a heat pump for simultaneous heating and cooling using HFC or CO 2 as a working fluid P. BYRNE (a), J. MIRIEL (a), Y. LENAT (a) Equipe MTRhéo Laboratoire LGCGM INSA de Rennes 20 avenue des buttes de Coësmes CS 14 315 35 043 Rennes Cedex France Tel: +33 2 23 23 42 97 Fax: +33 2 23 23 40 51 paul.byrne@univ-rennes1.fr ABSTRACT This article presents the design of a Heat Pump for Simultaneous heating and cooling (HPS) of hotels, luxury dwellings and smaller office buildings. It realizes simultaneously space heating and space cooling. It also preheats hot water for domestic use. The ambient air is used as a balancing source. During winter, some energy recovered by subcooling of the refrigerant is stored in the cold water tank. This energy is used to run a defrosting system and to improve the annual energy efficiency. Two refrigerants have been studied: R407C and CO 2. HFCs provide interesting performances, but new regulations on F-gases have led us to study low- GWP refrigerants. CO 2 seems to be a refrigerant which adapts well to the operation of the HPS. It enables direct DHW production and improves the defrosting system. Simulation results show that despite the 20% lower performance using CO 2, it offers good working conditions to the HPS. 1. CONTEXT AND OBJECTIVES Thermal performance of buildings is continuously improving and comfort requirements are more and more demanding. Whilst domestic hot water (DHW) demands continue to increase, there is also a rising demand for cooling to compensate internal heat gains caused by more and more household electrical equipment in buildings. A better thermal envelope implies that new buildings need far less energy for heating in the winter but that the indoor air temperatures can reach uncomfortable levels during the summer. Therefore new buildings need a better balance between heating and cooling over the year. During summer, some buildings need energy for DHW production and space cooling at the same time. Between winter and summer, glass fronted buildings can need, simultaneously, cooling in rooms facing south and heating in rooms facing north. A heat pump has a heating capacity at the condenser and a cooling capacity at the evaporator. This study presents the reasoning behind the design of a heat pump that can satisfy fluctuating needs, simultaneous or not, in heating and cooling. During winter, thermal energy demands are exclusively for space heating and DHW production. A certain category of heat pumps uses ambient air as a heat source. They are quite easy to install but when outside temperatures are lower, performance is poorer and frost forms more quickly on the air coil. Air-to-water heat pumps are not very popular with customers because they scarcely produce enough heat when heating demands are at their highest. The heat pump our team has designed heats and cools simultaneously. To do so, two separate circuits have been envisaged: one for heating and one for cooling. During winter, the cooling circuit isn t used. The second objective of this heat pump is thus to take advantage of the cooling circuit to store a quantity of energy at first and subsequently use it for defrosting and improving the average performance during winter. This heat pump for simultaneous heating and cooling (HPS) is designed either for hotels where DHW demands are high, glass fronted office buildings and luxury dwellings. The idea is not only to design an energy saving heat pump adapted to special types of buildings with the knowledge that building methods and ways of living are changing, it is also about studying the environmental impact of working fluids. The refrigerants used in new heat pumps are HFCs and natural fluids. HFCs provide good performances, but these fluids are classified by the Kyoto Protocol (1997) in the category of green house gases. That s why they are likely to be banned in the future like what is happening in mobile air-conditioning. Among natural fluids carbon dioxide seems one of the most adapted to residential applications or offices because it fulfils the requirements of low toxicity and flammability. According to Gustav Lorentzen (1994) (1995), author of "Revival of carbon dioxide as a refrigerant", CO 2 is the best natural refrigerant as a substitute for HFCs. Peter Nekså (2002) thinks that carbon dioxide is a very promising fluid for space heating applications. CO 2 technology is already commercialized in some applications such as domestic water heating (Kusakari, 2006).
Faced with this context, we have built numerical models and run simulations to evaluate the performance of a HPS and compared the results to those of a standard reversible heat pump. Also, the possibility of using carbon dioxide instead of a HFC has been examined from technical and environmental points of view. 2. DESIGN OF THE HEAT PUMP FOR SIMULTANEOUS HEATING AND COOLING 2.1. Three main heat exchangers The HPS has three main heat exchangers: a condenser (or a gas cooler for carbon dioxide) that heats a water tank for space heating and domestic hot water preheating; an evaporator that cools another water tank for space cooling; and a balancing coil that works either as a condenser or an evaporator on ambient air to adapt production to needs. The latter exchanger evacuates any exceeding cold produced in a heating mode and any exceeding heat produced in a cooling mode. 2.2. Three operating modes The functioning of the HPS is based upon three modes: a heating mode using the water condenser and the air coil as an evaporator; a dual mode preparing hot and cold water using the water heat exchangers (condenser and evaporator); and a cooling mode using the water evaporator and the air heat exchanger as a condenser. All modes include domestic hot water preheating. Commutation of modes by opening and closing on-off electronic gates is managed by a programmable controller. The HFC HPS (figure 1) is equipped with a condenser producing hot water for space heating. It is possible to prepare domestic hot water at 60 C with an auxiliary heater. The cold tank is connected to the water evaporator and to a subcooling heat exchanger for heat storage during the winter. This feature enables to use the cold water as a cold source during a certain time instead of cooler ambient air. It increases the low pressure and improves the overall performance of the HPS. If needed, this time sequence is used to carry out defrosting on the air evaporating coil. e Figure 1. HFC HPS Figure 2. CO 2 HPS The main difference between the R407C and the CO 2 machines resides in the heat exchanger technology used so that it can withstand operating pressures above 100 bars. In the HPS circuit (figure 2), the water condenser becomes a gas cooler The CO 2 part of the heat exchanger is made up of microchannels in order to handle high pressures. The gas cooler is divided into three sections. The top section is used for domestic hot water preparation at 55-60 C, the middle section is dedicated to the production of hot water for space heating at temperatures between 25 and 45 C and the lower section is connected to the cold water tank. The progressive temperature decrease in the gas cooler enables the production of hot water for domestic use as required for the HPS. For this reason, even if it provides poorer performance than standard cycles with HFCs, the CO2 transcritical cycle seems to suit the residential and office applications selected for the HPS. Also, the defrosting technique benefits from the high amount of energy recoverable by subcooling (about 30% of the total heating capacity available). However, CO2 technology is more in a research and development phase and needs more improvement before being really competitive with HFCs in space heating applications.
2.3. Defrosting technique During winter, the HPS alternates between heating mode and dual mode in order to use to a maximum the water evaporator instead of the air heat exchanger. This alternation enables to carry out defrosting on the frosted coil when it is not involved. Moreover, the average COP is improved because the cold source temperature is higher. In the heating mode, under cold outside air temperatures, the fins of the evaporator frost over. The cold tank is used to store some heat for a delayed defrosting sequence. The amount of heat needed is gained by subcooling the refrigerant until the cold tank temperature reaches 15 C. The air evaporator and the cold tank are designed so that the level of frosting doesn t come so high in a heating mode. The HPS then changes mode and operates in the dual mode. The regulation stops the fan of the air heat exchanger and switches the on-off electronic gates in order to use the water evaporator instead of the air evaporator. The defrosting system is then automatic: after passing through the water condenser for domestic hot water production and space heating, the refrigerant enters the air heat exchanger at a temperature around 30 C and is subcooled by exchanging heat with the frosted fins. At the same time, a heat pipe forms between the two evaporators. Part of the gas coming out of the water evaporator migrates towards the air evaporator where the temperatures and thus the pressures are lower. The gas exchanges heat with the frosted fins then condenses and goes back to the water evaporator by gravity. The cold water tank is designed to contain enough energy to run the dual mode as long as necessary to complete the defrosting. The sequence finishes in the dual mode until the cold tank temperature decreases back to 5 C. Then the controller switches the HPS back to the heating mode and so on. The major advantage of this operating sequence is to carry out defrosting without stopping or even degrading the heat production. Moreover, the cold source temperature being higher, the heat production is more efficient. And finally, the possibility of frequent defrosting without worrying about short working times of the compressor implies thinner frost thickness, better mean transfer coefficients and better overall performance of the HPS. The average efficiency under low temperatures should be improved compared to the performance of heat pumps that use hot gas or reversed cycle defrosting methods. 2.4. High pressure control As pressures and temperatures are linked during condensation, high pressure control enables to optimize the condensation temperature and thus the heating capacity. A bottle with a rubber bladder is used as a liquid receiver after condensation. On one side of the bladder there is refrigerant in a liquid phase and on the other side there is refrigerant in a gas phase. To fill up the bladder, the gas comes from the outlet of the compressor where the pressure is highest. To empty the bladder, the gas is sent back to the refrigeration circuit via the oil return line. The high pressure control system also enables the control of the volume of liquid in the bottle. The volume of liquid in the different condensers depends on the mode. In the heating mode, condensation occurs in the water condenser. In the cooling mode, condensation occurs in the outside air heat exchanger. The high pressure control enables to choose between the water condenser and the air condenser and to regulate the heating capacity. If the chosen mode is the heating mode, the bladder is filled up with gas at a pressure higher than the pressure of the liquid until the set pressure is reached. The bladder inflates and drives the liquid out of the receiver to fill up completely the air heat exchanger and the subcooling heat exchanger and a small part of the water condenser. If however the chosen mode is the cooling mode, then the condenser becomes the air heat exchanger. The set pressure is the lowest possible. The bladder is emptied by driving the gas out of the bottle into the oil return line. The refrigerant in a liquid phase is pumped out of the condensers into the bottle. This system is essential to ensure that the air coil and the subcooling heat exchanger are full of liquid and that the condensation is completed in the water condenser. Otherwise part of the refrigerant could condense in the subcooling heat exchanger and all the energy available from the condensation would not be transferred to the hot water tank. The proper functioning of the control system depends upon a special liquid receiver being designed with the main objective to limit as far as possible thermal transfer through the bladder which would provoke condensation in the gas part of the bottle and evaporation in the liquid part. For the moment, such a liquid receiver hasn t been designed. Therefore the pressure control system uses an external compressed air source instead of refrigerant in a gas phase.
3. NUMERICAL MODEL 3.1. Simulation Tools The numerical approach is carried out in four steps. A first simulation using TRNSYS software is run to obtain the heating, cooling and DHW demands of a standard hotel of 45 bedrooms. In parallel, the steady state performance in heating and cooling is calculated for each mode (heating, cooling and dual). The simulation tool used for this calculation is also TRNSYS. Every component of the machine is defined as a model and the refrigeration cycle is modelled as the connection of the different components by the input and output variables. The calculation is repeated until all the variables have converged. The heating sequence alternating heating and dual mode is then studied using Excel. Operating times for each mode are calculated during a frosting defrosting sequence. The last step consists in coupling the demands with the heating and cooling capacities of the heat pump depending on the outside air temperature. This simulation enables to calculate using Excel the daily times of operation for each working mode and to evaluate the average daily and annual performances. 3.2. Building model The building modelled is a hotel of 45 bedrooms where domestic hot water consumption levels are high. Thermal zones, wall materials, glazed surfaces, solar protections, ventilation, occupancy and lighting schedules were defined using the building module of TRNSYS so as to calculate the hotel s daily needs in heating and cooling. This simulation takes into account a weather data file that is representative of the weather in Paris over a year. 3.3. Heat pump models Four modes are simulated in steady state so as to calculate the heating and cooling capacities and the coefficients of performance: heating mode, cooling mode and dual mode with and without defrosting. Each mode establishes the performance of the cycle for ambient temperatures varying from -15 C to 15 C in heating and from 25 to 40 C in cooling. The set temperature for cooling is constant and equal to 5 C. The set temperature for space heating follows a heating curve depending on ambient temperature. Therefore the performance is linked to the ambient air temperature, even in the dual mode. 3.4. Heat exchanger models The HFC heat exchangers are designed using the LMTD method (Meunier et al., 2005). The heat transfer coefficients and the steady state capacities are calculated using the McAdams correlations (eq. (1), eq. (2) and eq. (3)) and the Lockart-Martinelli correlation (eq. (4)). The heating and cooling capacities are calculated using these heat transfer coefficients. Nusselt number during sensitive heat exchange: Heat transfer coefficient during condensation: 0,8 0.4 Nu = 0,023Re Pr if heated (1) 0,8 0.3 Nu = 0,023Re Pr if cooled (2) where 3 2 λ ρ g 2 F = µ 1/3 ( Re) 0. 4 0. 0077 F2 1/ 3 α = (3) µ Nusselt number during evaporation: where F and 0,83 = 1+ 1,925 X tt X tt 0,8 0,4 Nu = F.0,023Re Pr (4) 1 x = x 0,9 ρv ρl 0,5 µ l µ v 0,1
The CO 2 heat exchangers are divided into ten sections to take into account the variations of the thermodynamic properties of carbon dioxide according to temperature. The NTU and the LMTD calculation methods are used respectively in the gas coolers and the evaporators. The correlations used to calculate the heat transfer coefficients for carbon dioxide are the Gnielinski correlation (eq. (5)) in the gas cooler and the Lockart-Martinelli (eq. (4)) in the evaporator. To calculate the heat transfer coefficients for water, the McAdams correlations are used (eq. (1) and eq. (2)). where f = (0,79 ln(re) 1,64) 2 ( f / 8)( Re 1000) Pr Nu = (5) 1,07 + 12,7 2 2 / 3 ( f / 8) 1/ ( Pr 1) 3.5. Compressor model The compressor model calculates the mass flow rate and the compression work knowing the input and output pressures (eq. (6) and eq. (7)). Isentropic efficiency is assumed to be a constant equal to 75% for carbon dioxide and to 70% for R407C. These figures are mean values found respectively by Huff et al. (2003) and Pérez-Segarra et al. (2005) in their works on CO 2 and R134a that has properties comparable to R407C. W& = n 1 n m& P n 2 P 1 1 n 1 ρ P1 (6) W = m& (7) h cp 3.7. Operating strategy The coupling of the HPS with the building is a simulation based on a daily time step. For each day of the year, the heating and cooling demands are summed and the average heating and cooling capacities in each mode are calculated. The operating times per mode are obtained by dividing demands by capacities. The demands can be satisfied under several modes. The HPS starts to work in the dual mode as this mode is the most efficient (Byrne et al., 2006, 2007). Then it works in the heating or the cooling mode depending on the remaining thermal demands. Domestic hot water production isn t simulated in the same way for the HFC as for the CO 2 heat pump. For R407C, the DHW production is carried out by the heat pump up to 40 C and by an electric auxiliary heater from 40 to 55 C. For CO 2, all the DHW demand is satisfied by the the heat pump thanks to the progressive temperature decrease in the gas cooler. The daily and annual coefficients of performance are calculated in proportion to the operating times per mode. Electric consumptions are deduced from these daily calculations. They include the consumption of the auxiliary heater if needed. The calculations are also carried out for a standard reversible heat pump that has the same heating and cooling capacities as the HPS. 4. RESULTS AND ANALYSES 4.1. Building simulation The annual simulation of the 45-bedroom hotel (Figure 3) shows the variation of daily heating and cooling demands all year round. Annual needs amount to 25322 kwh for DHW production, 68640 kwh for space heating and 30146 kwh for space cooling. Paris is situated in a temperate climate where heating needs are higher than cooling needs. The HPS uses its high pressure control system in order to regulate in priority the heating capacity. It can also be pointed out that during spring and autumn heating and cooling demands appear simultaneously. During summer, the plateau in the heating curve corresponds to the domestic hot water demand that is nearly constant. From the end of April to the end of October, part of the hot and cold productions is simultaneous.
Needs in kwh 1400 1200 1000 800 600 400 Space Heating + Hot Water Production Cooling 200 0 0 50 100 150 200 250 300 350 days Figure 3. Daily needs in heating and cooling of a 45-bedroom hotel in Paris 4.2. Steady state energy efficiencies Figure 4 shows the coefficients of performance calculated in steady state simulations. The performance with carbon dioxide in heating and cooling is not as good as with R407C. The heating mode is 28% less efficient, the cooling mode is 43% less efficient and the dual mode efficiency is 31% lower. The difference in performance decreases as the outside temperature becomes more severe both in heating and cooling. In the heating mode, when the air temperature decreases, the performance of CO 2 tends to get closer to the performance of R134a. The same trend happens when the air temperature increases in the cooling mode. This phenomenon can be explained by the pressure ratios that vary more with R407C than with CO 2. The compression work depends directly on the pressures and HFCs are more sensitive to the outside temperature variations as the operating pressures are lower. Heating and Cooling Energy Efficiencies 8.00 7.00 R407C CO2-31% 6.00 5.00 4.00-28% -43% 3.00 2.00 1.00 0.00-15 C -10 C -5 C 0 C 5 C 10 C Heating mode 15 C Dual mode 25 C 30 C 35 C 40 C Cooling mode Outside air temperature Figure 4. Steady state energy efficiencies depending on temperature 4.3. Energy efficiencies Figure 5 clearly shows that CO 2 has a lower COP for space heating and cooling than R407C. The curves record a large increase during spring, summer and autumn, especially for R407C. Annual COPs are close to winter values because the operating times are higher during the heating period: 4.08 for R407C and 3.26 for carbon dioxide which corresponds to a difference of 20.1% over the year. R407C CO 2 10 Annual COP = 4.08 10 Annual COP = 3.26 9 9 8 8 Performance factor 7 6 5 4 3 Performance factor 7 6 5 4 3 2 2 1 1 0 0 0 50 100 150 200 250 300 350 0 50 100 150 200 250 300 350 days days Figure 5: Comparison of daily coefficients of performance between R407C and CO 2
4.4. Electric consumptions The electric consumptions of the HPS and of a standard reversible heat pump can be compared (figure 6). The defrosting sequence is supposed to take the same time for both machines. However, the standard heat pump that uses a reversed cycle for defrosting has to pump some heat out of the hot water tank to run its defrosting system. The model takes this parameter into account. The HPS working with R407C consumes 14% less electricity than a standard reversible heat pump. With carbon dioxide, the improvement reaches 22% thanks to the direct DHW production and to the favourable alternated sequence during the cold season. Part of the electric consumption used for the heating and cooling modes of the standard heat pump is replaced by a lower consumption in the dual mode of the HPS. The auxiliary heater consumption is also lower for the HPS. The dual mode is used a lot more with CO 2. It means that the dual mode is used principally for simultaneous demands in cooling and DHW production. The replacement of a standard heat pump by a HPS leads to a higher increase in performance with carbon dioxide than with a HFC. kwh per year 75 000 50 000 25 000 Electric consumption per mode for HFC and CO 2 heat pumps auxiliary heater for DHW dual mode heating mode 0 HFC HPS HFC Reversible heat pump CO2 HPS CO2 Reversible heat pump cooling mode Figure 6: Electric consumption in different modes, for a HPS and a reversible heat pump, and for R407C and carbon dioxide 5. CONCLUSION A Heat Pump for Simultaneous Heating and Cooling has been designed in the aim of heating, cooling and producing domestic hot water for luxury dwellings, hotels and smaller office buildings. This machine proposes an answer to reduce the performance loss of air-to-water heat pumps under low ambient temperatures and especially during defrosting sequences by alternation between an air evaporator and a water evaporator. The HPS has been thought up using two different working fluids. Today, HFCs are the most efficient and secure fluids on the market but they are classified as greenhouse gases, responsible for global warming. Carbon dioxide is environmentally friendly but it is less efficient as a working fluid. CO 2 technology needs more development to be really competitive in space heating applications (Kim et al., 2004). Nevertheless its thermodynamic properties enable domestic hot water production. Moreover, carbon dioxide offers a high amount of energy recoverable by subcooling that is profitable to the winter alternated operation of the HPS. These features have to be deeply explored for the CO 2 HPS. In order to compare HFCs and carbon dioxide, simulations have been run to calculate the steady state performance of each system. As expected, CO 2 offers lower performances than HFCs. The simulation of the installation of a HPS in a hotel of 45 bedrooms in Paris has then been studied. This heat pump can satisfy the needs in heating, cooling and domestic hot water production. The HPS records better annual performances compared to a standard reversible heat pump thanks to the simultaneous hot and cold productions and to the alternation between heating and dual modes during winter. The improvement is higher with carbon dioxide because of the direct domestic hot water production and the high subcooling capacity. Performance could be improved even more if simultaneous needs in heating and cooling were higher. For example in cities where climatic conditions are more severe than in Paris, such a heat pump would work more using the dual mode during winter sequences because of lower air temperatures and during spring and autumn because of greater daily temperature amplitudes. A R407C prototype is currently under testing. The modal operating strategy and the defrosting technique will be validated. Experimental results will help to formulate a more precise model and come closer to the real operation of this heat pump for simultaneous heating and cooling.
ACKNOWLEDGEMENTS The authors would like to acknowledge the Regional Council of Brittany (Conseil Régional de Bretagne) for providing financial support to this communication. NOMENCLATURE Letter symbols: CO 2 Carbon dioxide COP Coefficient of performance DHW Domestic hot water f Darcy-Weisbach friction factor (-) F Coefficient for Lockart-Martinelli parameter (-) F 2 Condensing coefficient factor (-) g Gravitational acceleration (m.s -2 ) GWP Global Warming Potential h Enthalpy (kj.kg -1 ) HFC Hydrofluorocarbon HPS Heat Pump for Simultaneous heating and cooling LMTD Log Mean Temperature Difference ( C) m& Mass flow rate (kg.s -1 ) n Polytropic coefficient (-) NTU Number of Transfer Units (-) Nu Nusselt number (-) P 1 Low pressure (bar) P 2 High pressure (bar) Pr Prandtl number (-) Re Reynolds number (-) x Vapor fraction (-) X tt Lockart-Martinelli parameter (-) W & Compression work (W) Greek symbols: α Heat tranfer coefficient (W.m -2.K -1 ) Variation λ Thermal conductivity (W.m -1.K -1 ) µ Viscosity (Pa.s) ρ Density (kg.m -3 ) Subscripts: cp Compression l Liquid v Vapor REFERENCES 1. Huff H.-J., Radermacher R. 2003, CO 2 compressor-expander analysis, HVAC&R Research for the 21 st Century program (21-CR), Air-Conditioning and Refrigeration Technology Institute, US, ARTI- 21CR/611-10060-01. 2. Kim M.-H., Pettersen J., Bullard C. W. 2004, Fundamental process and system design issues in CO 2 vapour compression systems, Progr. Energ. Combust. Sci. 30: 119-174. 3. Kusakari K. 2006, The spread situation and the future view of the CO 2 refrigerant heat pump water heater in Japan, 7th IIR Gustav Lorentzen Conference, Trondheim, keynote speech. 4. Lorentzen G. 1994, Revival of carbon dioxide as a refrigerant, Int. J. Refrig. 17(5): 292-301. 5. Lorentzen G. 1995, The use of natural refrigerants: a complete solution to the CFC/HCFC predicament, Int. J. Refrig. 18(3): 190-197. 6. Meunier F., Rivet P., Terrier M.-F. 2005, Froid Industriel, Editions Dunod: 20-42, 305-312. 7. Nekså P. 2002, CO 2 heat pump systems, Int. J. Refrig. 25(4): 421-427. 8. Pérez-Segarra C.D., Rigola J., Sòria M. 2005, Oliva A., Detailed thermodynamic characterization of hermetic reciprocating compressors R134a, Int. J. Refrig. 28(4): 579-593. 9. Protocole de Kyoto à la convention-cadre des nations unies sur les changements climatiques, 1997. 10. Byrne P., Miriel J., Lénat Y., Maré T. 2006, Optimization and modelling of a CO 2 heating cooling pump, Climamed: 3rd Mediterranean Congress of HVAC Engineering, Lyon, France, 379-388. 11. Byrne P., Miriel J., Lénat Y. 2007, Modelling of a heating and cooling plant coupled to a hotel using HFC or CO 2 as a working fluid. Comparison of the performances, Clima 2007 Wellbeing Indoors Congress, Helsinki, Finland, paper n 1234.