A closer look at [CO.sub.2] as a refrigerant.
Carbon Dioxide (R744) was a well known and widely accepted refrigerant in the early 1900's, but its popularity reduced with the introduction of fluorocarbons. The revival of R744 as a refrigerant started over a decade ago in Europe with the work of Dr. Gusav Lorentzen and Dr. Jostein Petterson (1). This sudden rediscovery was invoked by growing environmental concerns of global warming and ozone depletion.
R744 has some very attractive properties, which makes it destined to be used as a working fluid. It is non-flammable, non-ozone depleting, has good heat transfer properties, a high volumetric capacity, it is easily available and economic. However its critical temperature is 31.1[degrees]C, which is generally lower than the heat rejection temperature of a typical refrigeration and air conditioning system. Thus, wherever the heat rejection temperature is greater than the critical temperature, R744 must operate in a transcritical cycle, i.e, with a sub critical low-side pressure and a supercritical high side pressure.
The work of Dr. Peter Neksa has already proved the advantages of using R744 for water and space heating applications. In the field of automobile cooling systems, R744 also has proved advantageous over the conventional system in terms of better cooling performance, improved fuel consumption and zero ozone depletion rate (2). There is a drive to move R744 towards space cooling and it is being developed and tested across Europe.
The critical temperature is the temperature above which there is no clear distinction between liquid and gaseous phase. As the critical temperature is approached, the properties of the gas and liquid phases become the same. Above the critical temperature, there is only one phase (supercritical fluid) that is characterized by density and no latent heat effects. The critical pressure is the vapor pressure at the critical temperature. R744 has a critical pressure of 7.38 MPa at the critical temperature of 31.1[degrees]C.
In a normal refrigeration cycle, the gas from the compressor outlet is condensed in the condenser, by removing latent heat of condensation. But in a transcritical [CO.sub.2] cycle the discharge pressure of the compressor is above the critical point, where heat transfer cannot take place by phase change (condensation). In such a cycle, the gas from the compressor is cooled in a gas cooler, causing the density of the gas to increase, while temperature decreases. The supercritical temperature and pressure are not coupled, so they can be optimized, giving us an additional degree of freedom.
COMPARISON OF NATURAL REFRIGERANTS
In this section we compare the COP and volume flow rates of ammonia (R717), carbon dioxide (R744) and water (R718). It has already been derived that the thermodynamic and the transport properties of R744 are comparable with other refrigerants. (1)
In Figure 1, the x-axis represents the evaporator temperature and the y-axis, the temperature lift. Temperature lift is defined as the temperature difference between the evaporator and the condenser/gas cooler of a refrigeration system. The assumptions made for the comparative study are as follows:
[FIGURE 1 OMITTED]
a. Single stage compressor with polytropic efficiency of 100% was assumed in each case.
b. There is no superheating of the refrigerant.
c. There is no energy exchanged with the environment.
d. There is no subcooling of the refrigerant.
e. There is no pressure loss in the piping or the heat exchangers.
f. Optimized gas cooler pressure for R744
Figure 1 shows an operating range for the evaporator temperature from -55[degrees]C to 25[degrees]C for R744, as compared to -1[degrees]C to 265[degrees]C for R718, and -40[degrees]C to 105[degrees]C for R717. From the figure it is also possible to obtain the performance of the shown refrigerants for a particular condenser and evaporator temperature. As an example the performance of the various cycles for a condenser temperature of 35 [degrees]C versus varying evaporator temperature has been plotted using the dash-dot line.
Compared to R717 and R718, R744 always shows the best COP in the temperature range -55[degrees]C to -35[degrees]C, for the temperature lift of up to around 55 K. With decreasing temperature lift down to 15K, this range extends to an evaporator temperature of about 15[degrees]C. Therefore R744 can certainly be preferred for food freezing and cryogenic process industry and air-conditioning (AC) in moderate and northern regions.
While this study was conducted for the use of an expansion valve at the inlet to the evaporator, exergy studies performed for transcritical systems have shown that the COP of the cycle can be enhanced 33% by using an expander in place of the conventional throttling valve. (3)
The volume flow rate of R744 is considerably lower than that of other refrigerants (Figure 2), which makes it ideal for miniaturization. For an evaporator temperature of 20 [degrees]C, R744 has a volume flow rate approximately 10 times lower than the nearest competitor (R-22).
[FIGURE 2 OMITTED]
Studies comparing R744 with other refrigerants have shown that the risk of the high pressure in a R744 system is negated by the fact that the energy contained in the system is relatively lower than that in a R-22 system owing to the lower volume and refrigerant charge. (1)
The high-pressure low-flow rate also allows for the design of small diameter tubing or even micro-channel cooling. This property can well be used for micro applications like cooling of electronics and lasers systems. Due to the high pressures involved, the use of conventional material like silicon for the microchannels can be challenging, diamond microchannels can be an alternative. Diamond has good thermal conductivity and is already being used as a heat sink in micro chips. Using diamond, it is possible to attain high fin efficiency for heat sinks. (4) However, pressure tests are yet to be done on diamond micro-channels for pressure ranges encountered in the R744 system. Research done on stainless steel and aluminum micro-channels has shown that due to the reduced diameter they are able to withstand the high pressure. (4)
COEFFICIENT OF PERFORMANCE
R744 systems operates in two cycles, in a reversed Rankine cycle below the critical point and in a transcritical cycle, if the state of the discharge gas of the compressor is above the critical pressure. The enlarged iso-COP lines for R744 only, assuming the same condition as above are shown in Figure 3. All curves were obtained for optimized pressure ratio as discussed in the following section. Once the optimum pressure ratio is determined, the isentropic specific work done by the compressor can be calculated by
[FIGURE 3 OMITTED]
[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII]
[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] = the specific gas constant for [CO.sub.2]
[gamma] = the ratio of specific heats
[p.sub.2] = the discharge pressure of the compressor
[p.sub.1] = the suction pressure of the compressor
[T.sub.1]= the suction temperature of the compressor
The specific evaporator heat can be calculated from the specific enthalpy difference between the fluid entering and leaving the evaporator. The ratio of the specific evaporator heat to the specific compressor work gives the COP.
In the reversed Rankine cycle the compressor outlet pressure is dictated by the condensation pressure corresponding to the condenser temperature. In the transcritical cycle, the compressor outlet pressure can be varied to any value above the critical pressure. However, for a maximum COP there exists an optimum compressor discharge pressure corresponding to the gas cooler outlet temperature (Figure 4).
[FIGURE 4 OMITTED]
Since the cycle has no superheating, the evaporator temperature dictates the compressor suction pressure (Figure 5).
[FIGURE 5 OMITTED]
Combining the optimum pressure corresponding to the gas cooler temperature and the evaporator pressure, the compressor pressure ratio can be calculated (Figure 6). These results are consistent with those presented by Dr. Liao, Zhao and Kakobsen, which were limited to a very small temperature range. (5)
[FIGURE 6 OMITTED]
It is seen that although R744 operating pressure is much higher than that of other refrigerants, the pressure ratio of the compressor is comparatively small. Thus the R744 compressor can operate with considerably higher isentropic efficiency if designed and produced with same quality (polytropic efficiency). However, some studies conducted for R744 compressors have shown that the advantage of higher efficiency is often reduced by an larger entropy production in the compressor. (6)
For a better understanding of the effect of the pressure ratio on the COP, both iso-COP lines and iso-pressure lines have been plotted against the same axis, i.e, evaporator temperature vs. temperature lift in Figure 7.
[FIGURE 7 OMITTED]
In this section the effect of polytropic efficiency ([[eta].sub.poly]) of the compressor on the cycle performance is investigated. The polytropic efficiency can be related to the COP via the isentropic efficiency ([[eta].sub.is]) of the compressor. For this, the polytropic exponent (n) can be calculated with the isentropic exponent ([gamma]) for various polytropic efficiencies:
[[n - 1]/n] = [[[gamma] - 1]/[gamma]](1/[[eta].sub.poly])
Once we have the polytropic exponent, we can calculate the polytropic work done by the compressor:
[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII]
We can now calculate the isentropic efficiency of the compressor corresponding to the polytropic efficiency. The variation in isentropic efficiency is a function of the pressure ratio.
[[w.sub.is]/[[eta].sub.is]] = [[w.sub.poly]/[[eta].sub.poly]]
Using the definition of COP, we can calculate the COP incorporating the isentropic efficiency of the compressor. Figures 8 and 9 show the iso-COP lines for a fixed polytropic efficiency, indicating the variation of the isentropic efficiency with evaporator temperature as a function of the pressure ratio.
[FIGURE 8 OMITTED]
[FIGURE 9 OMITTED]
Figure 10 shows the iso-COP lines, the iso-pressure ratio lines and the isentropic efficiency lines corresponding for evaporator temperature vs. temperature lift considering 70% polytropic efficiency of the compressor.
[FIGURE 10 OMITTED]
The study presents COP maps for conventional and transcritical R744 ([CO.sub.2]) systems. It also provides the optimum pressure ratios for a wide range of gas cooler or condensation temperatures and evaporator temperatures as they have been used in this work. For certain compressor quality (polytropic efficiency) maps are generated that summarize the compressor requirements in terms of pressure ratio and isentropic efficiency, along with the obtainable COP for a wide range of combinations of evaporator temperature and temperature lift.
It can be concluded that R744 is a viable natural refrigerant, especially for lower temperatures or moderate temperature lifts. The volume flow rates of R744 are very small as are also the pressure ratios. This combination renders R744 also as an ideal refrigerant for micro-scale cooling systems.
Phillip Johnson, Director of Engineering, McQuay International, Staunton, VA: More than 100 years of industrial refrigeration industry practice agree with this paper's conclusion that carbon dioxide (R-744) is an excellent refrigerant for low-temperature refrigeration applications with moderate lift. The empirical evidence is the common use of carbon dioxide in the low-temperature side of cascade-type refrigeration systems, ** which tend to operate in the region of Figure 1 identified as "R744 best." While the title and focus of this paper is on carbon dioxide, an interesting observation is the good performance of water (R-718) at higher-temperature applications, as shown in Figure 1. Contrary to the case of carbon dioxide, the air-conditioning and refrigeration industry has not historically adopted water as a refrigerant. It would be interesting to see an analysis similar to Figure 1, comparing water to the refrigerants most commonly used by the industry. ** See the following references in the 2006 ASHRAE Handbook--Refrigeration: Chapter 3, pp. 3.26-3.27; Chapter 7, pp. 7.23-7.24; and Chapter 16, p. 16.7.
(1.) Man-Hoe Kim, Jostein Pettersen, Clark W. Bullard, 2003, Fundamental process and system design issues in [CO.sub.2] vapour compression system.
(2.) Peter Neksa, Jostein Petterson and Geir Skaugen, 2006, [CO.sub.2] Refrigeration, Air Conditioning and Heat Pump technology.
(3.) Jun Lan Yang, Yi Tai Ma, Min Xia Li and Hai Qing Guan, 2004, Exergy analysis of transcritical carbon dioxide refrigeration cycle with expander.
(4.) Kenneth Goodson, Katsuo Kurabayashi and R, Fabian W. Pease, 1997,Improved heat sinking for laser-diode arrays using micro-channels in CVD diamond.
(5.) Liao S.M, Zhao T.S. and Jakobsen A., 2000 A correlation of optimal heat rejection pressures in transcritical carbon dioxide cycles.
(6.) J.S. Brown, Y. Kim and P.A. Domanski, 2002, Evaluation of carbon dioxide as R-22 substitute for residential air-conditioning.
Norbert Muller, PhD
Norbert Muller is an assistant professor and Jijo Oommen Joseph is a master's student in the Department of Mechanical Engineering at Michigan State University, East Lansing, MI.
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|Title Annotation:||carbon dioxide|
|Author:||Muller, Norbert; Joseph, Jijo Oommen|
|Date:||Jul 1, 2009|
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