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The Need For Replacement Refrigerants

Since their development in 1931, chlorofluorocarbons (CFCs) were thought to be ideal refrigerants. They had chemical stability and relatively low toxicity, making them safe for both residential and industrial use. CFCs were also quite inexpensive, resulting both in a proliferation of refrigerators and air conditioners that utilized them, and in an unwillingness to repair systems with leaks, since they could often be periodically recharged for much less.

In 1974, CFCs were tentatively identified as destructive to the ozone layer (Domanski, 1997). For the next decade, this relationship was investigated, and a quantitative statement that tied CFCs to the depletion of ozone was released by the World Meteorological Organization and the United Nations Environment Programme (WMO/UNEP) in 1985. The Montreal Protocol (1987), which was agreed to by nearly one hundred and fifty countries, froze CFC consumption in 1989 and pledged to cut it in half by 1998. In 1992, the Copenhagen Amendments went even farther, and halted the production of CFCs in developed countries by 1996. The effects of these protocols, assuming international compliance, can be seen in Figure 1.1 (Ennis, 1994).

Figure 1.1: Results of International Protocols
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With CFCs scheduled to be phased out, hydrochlorofluorocarbons (most notably HCFC-22) gained in popularity. While their production levels were controlled by 1992, they were not scheduled for complete elimination. HCFCs were still damaging to the ozone layer, but much less so than CFCs. Their average stratospheric lifespan was much shorter, and not as many of them penetrated into the upper atmosphere. For example, CFC-12 has a stratospheric lifetime of 102 years, compared to a lifetime of 13.3 years for HCFC-22. Furthermore, the ozone depletion potential (ODP) of R-22 is only 5.5% of the ODP of R-12 (Sand et al., 1997). HCFCs and HCFC mixtures were developed that could serve as drop-in replacements for most of the CFCs in use. However, the excitement over HCFCs was short-lived, as the Vienna Convention of 1995 not only accelerated the HCFC-reduction timetable, but also required that their production effectively cease by 2020, with a complete cessation by 2030. Japan and some European countries have established cut-off dates that begin much earlier. In Switzerland, for example, HCFCs are banned by 2005.

Once again, replacement refrigerants need to be found, but this time there are no obvious solutions. While some single-component refrigerants present reduced-performance possibilities, the solution appears to lie with synthetic mixtures. These mixtures may be azeotropic, near-azeotropic, or zeotropic.

Mixture Properties


Figure 1.2: A Zeotropic Mixture of Ammonia and Water
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For a zeotropic mixture (also known as a non-azeotropic mixture), the concentrations of the liquid and the vapor phase are never equal (Alefeld & Radermacher, 1994). This creates a temperature glide during phase change (at which point the concentrations of the vapor and the liquid are continually changing). Zeotropic mixtures are the most common type of refrigerant blend. An example of a zeotropic mixture is ammonia and water. As can be seen in Figure 1.2, at no point do the bubble and dew point lines meet (except, of course, where there is pure ammonia or pure water). When the mixture is cooled, liquid begins to form at the dew point temperature, but this is not completed until the bubble point temperature.

This difference between the dew and bubble temperatures is known as a temperature glide. The smaller the temperature glide, the less loss of heat transfer due to concentration differences. Generally, zeotropic mixtures are not ideally suited to be placed in existing equipment, but can bring performance improvements when used with modified systems. Because of the varying liquid and vapor compositions, systems employing zeotropic refrigerant mixtures must be liquid charged. Doing otherwise could change the mixture's composition, which could result in decreased performance and increased safety risks over time.

Figure 1.3: Temperature vs. Enthalpy for Water
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Figure 1.4: Ammonia-Water Mixture, Concentration = 0.98
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Another feature common to zeotropic mixtures is a nonlinear temperature versus enthalpy profile. A single-component or azeotropic refrigerant will have a linear temperature profile, as in Figure 1.3, but a zeotropic mixture's profile can be strikingly nonlinear, as seen in Figure 1.4. This sort of temperature-enthalpy behavior results in a varying specific heat and raises the possibility of a temperature pinch within a heat exchanger, both of which complicate traditional heat exchanger calculations.


Figure 1.5: A Near-Azeotropic Mixture
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For a near-azeotropic mixture, the vapor and liquid concentrations at a given temperature and pressure differ only slightly. Most azeotropic refrigerant mixtures become near-azeotropic when the pressure or temperature is varied from the azeotrope point. R-410A (which is also known as AZ-20 under the AlliedSignal patent) is a near-azeotropic mixture of R-32 and R-125 (fifty-fifty mass percent). For standard condenser pressures and temperatures, the bubble and dew points for this concentration vary by less than 0.1 ° C (see Figure 1.5). Although it appears that this variation could be reduced even further by increasing the concentration of R-125, this is actually undesirable. For example, R-410A was developed to serve as a replacement for HCFC-22, and changing its composition reduces its suitability for that role while also increasing its flammability and toxicity. Near-azeotropic mixtures usually work fairly well with existing equipment.


An azeotrope is defined as a point at which the concentration of the liquid and the vapor phase is the same for a given temperature and pressure. Some mixtures have more than one azeotrope at a fixed pressure or temperature, but this is uncommon. At an azeotrope, a mixture behaves like a single-constituent system. Almost all azeotropic refrigerants have a boiling point lower than either of the constituents (which is known as a minimum temperature or maximum pressure azeotrope). An exception to this is R-507, which is a fifty-fifty weight percent blend of R-125 and R-143a, and is proposed as a replacement for R-502. A plot of temperature versus composition for a mixture of R-125 and R-143a can be found in Figure 1.6. While an azeotropic mixture may appear to be an obvious replacement for a pure refrigerant, there is no azeotropic mixture replacement for R-22, one of the most popular of the HCFCs, which duplicates its cooling capacity and pressures. Azeotropic refrigerant mixtures are uncommon, and it appears unlikely that new azeotropes will be found.
Figure 1.6: A Mixture With an Azeotrope At 41.18% R-125
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Heat Exchanger Analysis for Mixtures

In the analysis of heat exchangers, the Log Mean Temperature Difference (LMTD) is a standard calculation used to compute the required size (indicated by UA). The practice of using the LMTD for heat exchanger calculations dates to the 1950s (Chen, 1988). Other mean temperature methods, such as Underwood's (1933) one-third rule (qm1/3 = [1/2](q11/3 + q21/3) ), predate the LMTD, but none of them are as widely used today. The LMTD ceases to be valid, however, when a zeotropic mixture that exhibits a nonlinear temperature glide is used as the working fluid. Examining Figure 1.4, it is obvious that errors would result from using the LMTD to calculate the heat exchanger size for this mixture. While the inadequacy of the LMTD has been briefly commented on by Lundqvist (1995), he has only examined the differences between subcooling, condensing, and superheating, and ignored the nonlinearities in the two-phase region.

In order to ascertain the magnitude of the error, a numerical analysis must be performed. It is of interest to determine an alternative calculation that is applicable to zeotropes. The first step is to return to the definition of the mean temperature difference in a heat exchanger. Some assumptions that are made in the formulation of the LMTD are acceptable: for example, that the overall heat transfer coefficient will remain nearly constant over the heat exchanger area. Obviously, though, it cannot be assumed that the temperature will be an easily-determined function of the heat exchanged. As a result, the integral evaluation of the variance in temperature between the two streams over a differential amount of heat transferred must instead be transformed into a summation.

Since the accuracy of this method will increase with the number of steps that are taken, the step size will be made sufficiently small so as to minimize numerical error. The UA that is found using the traditional LMTD can be contrasted with that found using the numerical method, and the deviations compared for various possible refrigerant replacements over a range of pinch points.

next up previous contents
Next: Derivation of the Log Up: Heat Exchanger Mean Temperature Previous: Abstract

Laura Atkinson Schaefer