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Description
This paper contains a review of the latest research activities toward transcritical C[O.sub.2] cycle technology as a replacement for the fluorocarbon-based vapor-compression technology for heat pumping, air-conditioning, and refrigeration applications. In addition, the potential of using the transcritical C[O.sub.2] cycle technology for particular applications and the barriers that must be overcome before its acceptance are reviewed. Advances in theory and recent experimental data are also included.
INTRODUCTION
In the early twentieth century, carbon dioxide (C[O.sub.2]) was extensively used as the refrigerant for vapor-compression refrigeration systems, especially in marine systems, due to its nonflammability and acceptable toxicity. However, C[O.sub.2] has unfavorable characteristics of lower critical temperature and higher operating pressure compared to ammonia ([NH.sub.3]) and sulfur dioxide ([SO.sub.2]). After the advent of chlorofluorocarbons (CFCs) in 1928, the use and interest in carbon dioxide as a refrigerant significantly diminished until the revival of interest in natural working fluids in the early 1990s due to the environmental concerns of global warming. Table 1 shows a comparison of the direct global warming potentials among some fluorocarbon-based refrigerants and C[O.sub.2].
A paper by Lorentzen and Pettersen (1992) appears to be the first open literature about revisiting the use of C[O.sub.2] as a viable alternative refrigerant to deal with global warming and ozone depletion. Since then, the transcritical cycle technology using C[O.sub.2] as the refrigerant has been the subject of increased research activities and is considered a possible replacement for vapor-compression cycle technologies that use fluorocarbon-based refrigerants in certain applications.
One of the prominent applications is automotive air conditioning. By now, most of the automobile manufacturers have prototype systems of C[O.sub.2] air conditioners. Several new and innovative designs for heat exchangers, compressors, and valves have emerged from studies in this area. This application has recently received increased attention based on the European Parliament's July 2006 vote to phase out the refrigerant HFC-134a, which is currently used in automotive air conditioning. This vote marks the first international action to ban a hydrofluorocarbon (HFC) refrigerant due to its high global warming potential. HFC refrigerants were introduced in response to the Montreal Protocol to phase out chlorine-containing refrigerants to protect the Earth's ozone layer. Another application for C[O.sub.2] is environmental control units (ECUs), which are packaged air-to-air air conditioners that are used in the cooling of mission critical electronics and personnel. The US Army currently maintains roughly 22,000 ECUs of varying capacity, either in service, in storage, or on order, that use HCFC-22 as their refrigerant. These need to be replaced by 2010, as this is the current phase-out date for HCFC-22 in new equipment due to the refrigerant's ozone depletion potential. A third application that shows great promise for transcritical C[O.sub.2] systems is heat pump water heaters. In fact, the first commercial product has already been introduced on the Japanese market. Finally, transcritical C[O.sub.2] technology has been proposed as the cooling system in glass door coolers and vending machines based on the interest by a leading soft drink manufacturer to use "green" cooling technologies.
This paper presents a description of the C[O.sub.2] systems that are proposed for each of these applications and the performance results obtained with these systems. Both theoretical modeling results and experimental data will be presented. Whenever feasible, a comparison to the existing fluorocarbon-based technology will be given.
BASIC TRANSCRITICAL C[O.sub.2] CYCLE
The critical point of C[O.sub.2] is 30.85[degrees]C and 73.53 bars. Thus, many refrigeration and air-conditioning applications span the critical point in that the heat absorption temperature is below and the heat rejection temperature is above the critical temperature. This implies a transcritical cycle in which the evaporator operates as a familiar vapor-liquid two-phase device but the condenser is replaced by a supercritical heat rejection device called a gas cooler. Figure 1 shows the cycle state points indicated in a p-h diagram. The transcritical C[O.sub.2] cycles were drawn based on the following assumptions: gas cooler outlet temperature of 30[degrees]C, evaporation temperature of 0[degrees]C, superheat of K, and isentropic compression and isenthalpic expansion processes.
[FIGURE 1 OMITTED]
The cycle shown in Figure 1 illustrates the basic transcritical cycle. Assuming a constant gas cooler exit temperature, the cycle can be operated at different gas cooler pressures by using the expansion valve to adjust the back (high-side) pressure. Varying the gas cooler pressure has two impacts on the cycle performance. As the gas cooler pressure increases, the enthalpy difference across the evaporator (cooling capacity) increases. At the same time, the enthalpy difference across the compressor (compressor work input) increases. Since these impacts have opposing effects on cycle performance, the efficiency of the transcritical C[O.sub.2] cycle, typically indicated by its coefficient of performance (COP), can be maximized by adjusting the gas cooler pressure. Figure 2 presents the COP as a function of gas cooler pressure for several evaporation temperatures and gas cooler exit temperatures. Figure 2 was generated by conducting a thermodynamic cycle analysis in EES (Klein 2004) using the assumptions on isentropic compression and isenthalpic expansion processes and a superheat of K.
Figure 2 indicates that both the gas cooler exit temperature and the evaporation temperature have significant impacts on the maximum COP. It can also be seen from Figure 2 that the discharge pressure should be in the order of 90-100 bars or higher depending on the application. This means that when the evaporation temperature is, for instance, 0[degrees]C, the discharge temperature in single-stage compression with dry saturated suction vapor will be about 70[degrees]C-80[degrees]C. This temperature can be adjusted by varying the discharge pressure and the suction vapor state using a suction-to-liquid line heat exchanger or, possibly, some liquid injection. A suction-to-liquid line heat exchanger, as used quite commonly in refrigeration and air-conditioning equipment, may also result in a 5%-10% increase in cycle efficiency (Lorentzen and Pettersen 1993; Robinson and Groll 1998a).
[FIGURE 2 OMITTED]
TRANSCRITICAL C[O.sub.2] CYCLE MODIFICATIONS
The basic transcritical C[O.sub.2] cycle is thermodynamically less efficient compared to the conventional vapor-compression cycle if one assumes that the evaporation temperatures are the same and that the gas cooler exit temperature is equal to the condensing temperature. However, several studies in the literature indicate that these assumptions are incorrect. Due to the superior heat transfer characteristics of C[O.sub.2] compared to halocarbon refrigerants, the gas cooler exit temperature of a transcritical C[O.sub.2] system is significantly closer to the heat sink temperature than the condensing temperature of conventional vapor-compression systems for similar heat exchanger designs (Robinson and Groll 2000). While the differences in the evaporator approach temperature are not as large as the ones on the heat rejection side, studies also indicate that the evaporation temperature of a C[O.sub.2] system is closer to the heat source temperature than the evaporation temperature of halocarbon systems for similar heat exchanger designs (Robinson and Groll 2000). Even considering better heat exchanger approach temperatures, the basic transcritical C[O.sub.2] cycle may still be less efficient than the conventional vapor-compression cycle; thus, several advanced cycle designs have been considered that show promise well beyond the basic cycle.
Two-Stage Compression
The compression power consumption may be reduced by utilizing technologies such as multiple-effect compression (e.g., a single compressor with economizer) (Voorhees 1905) or two-stage compression (Hall 1889) that were already available for C[O.sub.2] vapor-compression cycles at the beginning of the twentieth century. In addition, these technologies can be applied to increase the overall temperature lift between the lowest heat source temperature and the highest heat sink temperature, resulting in a major expansion of possible applications. A two-stage cycle with intercooling is shown in Figure 3. The intercooler and gas cooler can be implemented in one heat exchanger. The transcritical two-stage cycle with intercooling provides operating characteristics that are of great importance to the cycle efficiency. Figure 4 presents two different two-stage transcritical C[O.sub.2] cycles with intercooling in a pressure-enthalpy diagram. Both cycles have an evaporation temperature of 0[degrees]C, a gas cooler exit temperature of 35[degrees]C, a gas cooler pressure of 10 MPa, and isentropic compression and isenthalpic expansion processes. The cycles differ in their choice of intermediate pressure, i.e., 7 versus 8 MPa. It can be seen from Figure 4 that the choice of intermediate pressure has a significant impact on the work input of the two compressors.
[FIGURE 3 OMITTED]
It can also be seen from Figure 4 that intercooling is only possible when the intermediate pressure is high enough so that the first-stage compressor discharge temperature is higher than the outdoor, i.e., gas cooler, exit temperature. For lower intermediate pressures, the two-stage compression process equals a single-stage compression process and no benefit in cycle efficiency can be achieved.
Figure 5 presents the COP of the two-stage transcritical C[O.sub.2] cycle with intercooling as a function of intermediate pressure for the same operating conditions as used in Figure 4. The COP reaches a maximum at a relatively high intermediate pressure. This is a substantial difference from subcritical two-stage cycles with intercooling, where the intermediate pressure is typically chosen as the square root of the high-side times the low-side pressure. To explain the COP behavior as a function of intermediate pressure, Figure 4 needs to be considered. It can be seen from Figure 4 that as the intermediate pressure increases above the critical pressure, the second-stage compression process moves quickly into the supercritical region, where the slopes of the isentropic lines are steeper than in the conventional superheated region. In fact, the inlet state of the second-stage compression process moves to lower enthalpies than the critical enthalpy due to a slope of the isotherms of nearly zero in the supercritical region. As a result, the enthalpy increase through the second-stage compressor decreases sharply as the intermediate pressure rises across the critical pressure. Just before the slope of the isotherms changes again toward lower enthalpies, the total compression work reaches a minimum and, thus, the COP reaches a maximum. At maximum COP, the enthalpy difference across the intercooler (at intermediate pressure) is significantly larger than the enthalpy difference... |
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