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issue: March 2005 APPLIANCE Magazine

Engineering: Heating and Air-Conditioning
Residential Space Conditioning and Water Heating with Transcritical CO2 Refrigeration Cycle

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by Clark Bullard and John Rajan, University of Illinois, and Sung-Oug Cho, Samsung Electronics Co. Ltd.

The use of CO2 refrigerant for simultaneous space conditioning and water heating has been the focus of renewed research and testing since Lorentzen re-introduced the idea of the transcritical cycle [3]. Neska et al have investigated several CO2 systems under standard test conditions, including solo heat pump water heaters and dual-purpose systems that simultaneously provide space conditioning and domestic water heating [5]. Several authors have also published results of prototype CO2 systems operating over wide ranges of test conditions.

Figure 1. “Ideal” Simultaneous Air-Conditioning and Water Heating Cycle

This paper presents an idealized residential system that meets domestic water and space conditioning demands (see Figure 1). The ideal system presented heats water to U.S. domestic storage temperatures of 60°C, in addition to meeting residential heating and cooling loads. Using simplified thermodynamic calculations and assumptions, optimal system operation strategies and configurations were identified for both heating and cooling seasons. The results provide a basis for more detailed system simulations that can focus on optimizing actual heat exchanger hardware design.

The optimal operating strategy varies throughout the year because the transcritical CO2 cycle exhibits a COP-maximizing discharge pressure, with a discharge temperature above or below the desired hot water temperature. Optimization algorithms are used to identify the most efficient operating condition. Water heating costs can be reduced to a third of electric resistance at Tamb > -5°C, and the efficiency increases with Tamb. After reviewing the assumptions made, an overview will be given of the various operation strategies and hardware configurations selected by climate season. Each operating strategy based on ambient climate will be examined in detail. Summaries of system parameters, such as high and low end CO2 pressures and temperatures, will provide ideal system parameters that provide the starting point for more detailed system simulation and control studies of the space-conditioning and water-heating system.

Figure 2. Space-Conditioning and Water Heating Modes


A number of assumptions were made in order to greatly simplify cycle calculations and the optimizations, which were performed using Engineering Equation Solver (EES). Heating and cooling loads were based on an overall UA value from a moderately insulated house, approximated as linearly dependent on ambient temperature [6]. The system provides 10.5 kW (3 tons) of cooling at the standard test condition, Tamb = 35°C. The heating and cooling lines meet at 18°C, and it is assumed that ventilation suffices between 15°C and 20°C. A minimum water heating capacity of 4 kW is desired, matching common electric resistance heaters that meet the 12.1 kWh/day average U.S. domestic hot water demand in 3 hr. The compressor was sized for maximum displacement at Tamb = 40°C, with a speed range of 30 < Wc < 120 Hz. The isentropic efficiency is taken from a 10.5 kW Dorin compressor and expressed as a function of Pdis/Psuc.

Assumptions for the cooling season are summarized in Figure 2, which shows a CO2 T-h diagram for simultaneous air conditioning (a/c) and water heating (WH). A pinch limit, TCO2  > TH2O or air + 2°C, was enforced for the counter-flow water and indoor air heat exchangers. For the air-conditioning season, the Tevap is maintained at 13.5°C by modulating blower speed to meet sensible heat ratio target of 0.75 [1]. The outdoor coil is cross-flow and assumed large enough to cool CO2 to a 2°C approach at any ambient air temperature. An internal heat exchanger (IHX), assumed to have 0.9 effectiveness, proves beneficial as cooling loads increase. Pressure drop is neglected in all heat exchangers.

A simplified heat-pump cycle can be used for space heating (SH) or water heating. During heating season, UA = 3.6 was assumed for the outdoor coil, based on an airflow rate and geometry from a previously simulated coil [1]. Space-heating air is supplied to the room at 40°C and returned at 20°C. Water is heated from 17°C supply to 60°C delivery temperatures.

It is recommended that the IHX be bypassed in the heat-pump cycle because the highly superheated suction gas produces a higher compressor discharge temperature that could risk compressor oil breakdown; the impact on COP is negligible.

Figure 3. Water Heating Costs

Overview of Optimized System

Five distinct operating strategies are needed for this optimized system, as will be detailed later in Figure 5. The first two strategies apply to the a/c and WH cycle. For Tamb > 23.5°C, indoor air is the sole heat source for water heating. The water inlet valve to the heater is simply opened for simultaneous air-conditioning and water heating, and closed during a/c-only operation. At 20 < Tamb < 23.5°C, insufficient heat is available in the indoor air; therefore, a combination of indoor and outdoor air heat sources may be used. Thus, a third operation strategy is needed, using the CO2 system as a water-only heat pump during a/c off-cycle time. The system continues to act as a water heat pump during the ventilation season, 15 < Tamb < 20°C. In the fourth strategy, tandem operation between space heating and water heating cycles is employed when -4 < Tamb < 15°C. At Tamb < -4°C, the system reaches maximum capacity, and space heating loads must be met using a fifth strategy, providing water heat and supplementary space heat with an electric resistance heater.

Adding water heating to the overall CO2 system reduces the water heating cost by 2/3 or more compared to the 12 kWh/day needed by electric resistance. Figure 3 shows the cost of adding water heating load to the CO2 refrigerant system; the incremental COP ranges from 3 to 16. Cost declines with rising evaporating temperature as the outdoor temperature increases. One-stage water heating was found to be optimal or nearly optimal for all seasons, with the CO2 -to-H2O heat exchanger placed immediately downstream of the compressor. Where the water source is colder, it may be cost-effective to add a pre-heater because a colder heat sink increases cycle efficiency. Water heating times are influenced by compressor speeds, but are always faster than electric resistance.

At Tamb > 29°C (Pdis > 75 bar), there is abundant heat rejection to the air due to the sizable a/c loads, and the COP-maximizing discharge pressure for the a/c only cycle is high enough to heat water to 60°C. Therefore, water heating is free. The 3-hr water heating time constraint determines how much of the high-side refrigerant temperature profile is needed for water heating (4 kW). The rest of the heat is rejected to air in the gas cooler downstream. As a/c load decreases with Tamb, the COP-maximizing high side pressure decreases. When it reaches 75 bar at 29°C, the temperature difference between CO2 and water approaches the pinching limit and water heating is still free because no extra compressor work is needed. CO2 flows through the refrigerant side of the water heater continuously, while water flows through it for 3 hr per day and is then shut off for the remaining 21 hr. Rejecting all heat to the water would result in much faster heating times, but it is less efficient because the high-side discharge pressure would be 94.4 bar, regardless of Tamb, which is inefficient.

At Tamb < 29°C, the COP-maximizing discharge pressure at these moderate-load conditions is sufficient for a/c-only, but too low to heat water to 60°C. Therefore, added compression energy is required in order to limit water-heating time to 3 hr. At 26°C, the compressor speed reaches its (lubrication limited) minimum of 30 Hz. Compressor cycling occurs at Tamb < 26°C, meeting the daily cooling load in less than 24 hr. At Tamb = 23.5°C, the air-conditioning and water-heating system runs only 3 hr per day, with all 4 kW of heat rejection going to water heating. As ambient temperature drops and less heat is being removed from the building, the water-heating-only system compensates for the shortfall of the indoor air energy source.

When Tamb < 20°C, the outdoor air is the sole heat source for heating water and air. During the ventilation range, 15 < Tamb < 20°C, the CO2 system is used only as a water heat pump. When space heating demands start at Tamb < 15°C, water heating at 30 Hz may be performed in tandem with space heating, because heating loads are small at mild temperatures, off-cycle time is abundant. The space heating and water heating (30 Hz) times total 24 hr at 10.2°C.

At -4 < Tamb < 10.2°C, the space and water heating system configurations operate continuously in tandem at a common minimum compressor speed until the compressor reaches the maximum 120-Hz speed limit at around -4°C. Consequently, as compressor speed increases to meet higher space-heating demands on increasingly colder days, the water heating time will decrease.

The cycle state points are dictated by the compressor speed, which determines water and space heating capacity. While the compressor runs at constant speed, the transitions are accomplished by the back-pressure valve adjusting Pdis from approximately 75 to 90 bar in order to maximize COP for space heating and water heating, respectively. Suction and discharge pressures are shown in Figure 4.

Figure 4. Discharge and Suction Pressures

Choosing Optimal System Configurations

The total cost of water heating using one- or two-stage water heating strategies is very similar. For two-stage water heating, in which refrigerant flows through a pre-heater downstream of the outdoor coil, system efficiency is increased by the presence of this additional heat sink, saving about 1 kWh/day at Tamb > 28°C. Only in very hot climates would this small benefit offset the cost and complexity of the extra refrigerant circuit through the water heater. Benefits would also be larger in cities where water supply temperature is lower or hot water capacity demands are large. During cooler ambient conditions, there is very little heat rejection to preheat sink because the CO2 mass flow decreases along with the temperature difference between the water inlet temperature and that of the CO2 exiting the outdoor coil.

For the heating season, a simultaneous one-stage space heating and water heating strategy was compared to a tandem space heating and water heating operation and was found to save only about 0.5 kWh/day. Tandem mode is more efficient because the discharge pressure can be optimized for space heating and water heating independently.

Summary and Conclusions

Based on the preceding cycle analyses, it is apparent that a complex system configuration is not required to approach ideal water heating efficiency while satisfying residential a/c and heating loads. A rather simple single stage water heater will suffice, due to the mild 17°C city water heat sink. Therefore, future hardware-specific simulation analyses can be focused on this simple configuration to refine component and system designs. Conservative approaches to preventing compressor oil breakdown necessitate bypassing the IHX during the heating season because of high discharge temperatures.

The optimal operating and control strategies appear to be more complex, but this is partly an artifact of the need to make specific assumptions for initial cycle analysis. In practice, the selection of sensors will determine the complexity of the control system.

(CLICK HERE to see Figure 5.)


[1] Kim, M-H, Samsung Electronics Corporation, personal communication, Jan. 2002.

[2] Klein, S.A., Alvarado, F.L, Engineering Equation Solver, Version 6.881-3D (software), F-Chart Software, 2003.

[3] Lorentzen, G., 1994. Revival of Carbon-Dioxide as a Refrigerant, Int. J Refrig., Vol. 17, No. 1: pp. 292-301.

[4] Neska, P., 1998, CO2-Heat Pump Water Heater: Characteristics, System Design and Experimental Results, Int. J Refrig., Vol. 21, No. 3: pp. 172-179.

[5] Neska, P., 2002, CO2 Heat Pump Systems, Int. J Refrig., Vol. 25, No. 4: pp. 421-427.

[6] Richter, M. R., C. W. Bullard, and P. S. Hrnjak, 2001, Effect of Comfort Constraints on Cycle Efficiencies, Proc. International Mechanical Engineering Congress and Exposition, New York, NY, U.S.

This is an edited version of a paper originally presented at the 2004 International Refrigeration and Air Conditioning Conference held at Purdue, West Lafayette, IN, U.S.

About the Authors

  • Clark W. Bullard is a research professor of Mechanical Engineering at the University of Illinois at Urbana-Champaign. He founded the Air-Conditioning and Refrigeration Center, a 30-company NSF Industry-University Cooperative Research at the University of Illinois.
  • John Rajan is a graduate student of Mechanical Engineering at University of Illinois at Urbana-Champaign, where he will graduate with a Master's Degree in May 2005. He currently works at General Electric Power Gas Turbines in Greenville, SC, U.S. as an air-thermal systems design engineer.
  • Sung-Oug Cho, Ph. D is a senior engineer at the R&D Center, Digital Appliance Network Business of Samsung Electronics Co. Ltd. in Korea. He is a member of the Society Air-Conditioning & Refrigerating Engineers of Korea and is also active in the Korean Society of Mechanical Engineers.

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