Two-Stage Compressors In Environmental Test Chambers

Mechanically refrigerated test chambers are classified as either single stage low (-40 °C) or cascade ultralow (-70 °C) temperature systems. While cascade systems are reliable and versatile, the potential for leakage between stages within a cascade heat exchanger remains. Further, the need for two compressors and increased control mechanisms can make the system more and more complex. The only mechanical alternative to a cascade system that is utilized in most environmental test chambers is a single stage system. There are drawbacks to a single stage system however. While a single stage system may be more reliable, it cannot reach the ultra-low temperatures that a cascade system can primarily due to the large compression ratios involved and the diminished capacity at pressures below atmospheric. For those users who prefer the simplicity and reliability of a single stage system, yet desire increased capacity at -40 °C and/or have the desire to reach temperatures below -40 °C on an infrequent basis, a two-stage compression process employing a single refrigerant to reach the ultra-low temperature of a cascade system while at the same time providing a slightly increased capacity over that of a single stage system at -40 °C may be a solution an approach that hasn’t been utilized heavily within the industry. NOMENCLATURE dis = Discharge ρsuc = Density (kg/m) h = Enthalpy (kJ/kg) out = Outlet isen = Isentropic m& = Ref. Mass Flow Rate (kg/s) No. = Number Pwr = Power ref = Refrigerant RPM = Compressor speed (rev/min) suc = Suction η = Efficiency TXV = Thermostatic Expansion Valve V = Volume (m) cyl = Cylinders comp = Compressor eff. = Effectiveness INTRODUCTION Environmental test chambers began to enter the market in the 1960’s and since that time, they have been used for many different purposes in a myriad of different industries. Since the introduction of the basic cascade system in test chambers, manufacturers have built systems that employ many different types of refrigeration systems. Some of these include liquid nitrogen (or carbon dioxide) spray (or coil) and single stage systems. For ultra-low temperature applications (to -70 °C) however, cascade systems are usually the system of choice (versus a nitrogen system). The cascade system (Figure 1 below) utilizes two single stage compressors. In this system, the higher temperature system’s (R-404A or R-507) evaporator acts as the condenser for the lower temperature system (R-23 or R-508B). By utilizing two independent systems, it is possible to avoid large compression ratios while maintaining good capacity at low temperature (although this is dependent on the properties of the refrigerant being used). However, in the case of a cascade heat exchanger failure, the refrigerants would mix necessitating pumpdown of both systems, disposal of a mixed refrigerant, and repair of the heat exchanger. A single stage system (also shown in Figure 1 below) uses less components, and a solitary refrigerant, but is typically rated only to temperatures of -40 °C due to compression ratio and capacity issues inherent to the properties of the refrigerant used, typically R-404A. Figure 1: Typical single stage and cascade refrigeration systems utilized in test chambers. One system not typically utilized in environmental chambers is one that makes use of a two-stage compressor. With a two-stage compression system, it is possible to avoid the penalties associated with the use of a high compression ratio by using either direct injection, or an intercooler at the intermediate pressure level prior to the second compression process. COMPRESSOR SELECTION AND CHARACTERISTICS In order to validate the performance of a two-stage system and to make a comparison against currently available equipment, a two-stage compressor whose operating characteristics were available was chosen for use in this investigation. Although available in sizes from 5 to 60 hp (3.7 to 44.7 kW), a 5 and 8 hp (5.9 kW) model were chosen for use in simulation as these are sizes that are typically utilized on standard environmental test chambers. Figure 2: Curve fits for 5 hp two-stage compressor based on suction and discharge pressures. With data, the simulations were then able to include calculations for compressor power and intermediate pressure that were determined by calculating a three-dimensional surface using curves fitted from tabular data based upon the first stage suction and second stage discharge pressures (curves for the 5 hp model are shown in Figure 2 above). Volumetric efficiency curves were also plotted and compiled to form a surface, as measured values for mass flow rate in the compressor were also available in a tabular format, in addition to the cylinder displacement. A constant isentropic efficiency of 0.7 was assumed, which can said to be average for most reciprocating compressors. More accuracy in the results would be possible if further information was available, but as the selection will affect all simulations in the same manner, the constant value will still allow for comparison. Compression efficiency would follow from this setting of isentropic efficiency, but is not an important value in this study. ( ) comp ref suc isen dis isen Pwr m h h & * , − = η eq. 1 cyl suc cyl ref vol No RPM V m . * 60 ρ η ⋅ ⋅ = & eq. 2 Finally, although this compressor is rated for use with R-22, R-404A, or R-507, R-22 was not utilized in this study, as its finite lifetime and HCFC designation make it an unpopular choice in the construction of new environmental test chamber equipment. R-404A was chosen over R-507 as it is more commonly used on single stage systems within the industry, and therefore a good comparison could be made later against single stage systems. SIMULATION Using the formulas for the three-dimensional surfaces that outline the compressor’s performance, three EES simulations were written in order to determine the performance of the two-stage compressor in: a system with no intermediate cooling, a system with intercooling of differing percentages (10-40% of the overall mass flow rate), and finally, a system that utilizes direct injection cooling. The Mollier charts detailing these last two options are shown in Figure 3 below. Figure 3: Two-stage cycle with injection (left) and intercooling (right) circuits plotted on Mollier charts. Cycle with No Second Stage Cooling Initially, the two-stage 8 hp compressor simulation utilized no intercooling or injection (in order to determine baseline performance). The simulation (and all of the simulations that follow) made the assumption that there was no heat transfer to or from the compressor, and that there was no sub-cooling at the condenser outlet (saturated liquid). For a condenser temperature of 32 °C, an evaporator temperature ranging from -15 to -75 °C, and a superheat of 20 K (the system will adjust itself to prevent excessive discharge temperature by limiting the temperature difference between the coil and chamber workspace), the second stage discharge temperature and evaporator capacity ranged from 69 to 126 °C and 16.6 to 0.53 kW respectively. The 5 hp compressor’s capacity ranged from 11.63 to 0.36 kW. The corresponding values for COP ranged from 2.7 through a 77% drop to 0.62 for the 5 hp compressor. This is to be expected as the efficiency will decrease with the increased compression ratio. Results for the 8 hp unit are shown below with the 5 hp unit in Figure 4. 0 2 4 6 8 10 12 14 16 18 0 10 20 30 40 50 60 70 80 90 Compression Ratio C ap ac ity (k W ) 0.0 0.5 1.0 1.5 2.0 2.5 3.0 C O P 8 hp Compressor 5 hp Compressor Open/Closed Symbols: COP/Capacity Figure 4: Capacity and COP of the 5 and 8 hp compressors in a cycle with no intermediate cooling. Cycle Utilizing Injection Cooling The mildly excessive final discharge temperature (more so at -75 °C, chambers are usually limited to 116 °C) and lack of capacity and COP at low temperature demonstrates the need for cooling of the first stage discharge in order to prevent compressor damage and increase performance. One option previously mentioned is to throttle the refrigerant at the condenser outlet to the system’s intermediate pressure level, and separate the saturated liquid and vapor that results. The saturated liquid at the intermediate pressure level is then throttled again and passed through the evaporator (increasing the enthalpy difference), while the vapor is mixed with the first stage discharge gas in order to cool it before entering the second stage compression process. The mass flow rate in the system stays constant (the injected mass flow rate is pulled from a vapor tank and is not passed along to the system) in this simulation, while the enthalpy difference in the evaporator increases yielding higher capacity and COP as shown in Figure 5 below. Not only is the performance of the system enhanced by using an injection strategy, but the final discharge temperature is lowered by 8 to 32 K, while the capacity is also raised by 30 to 100% (all for evaporator temperatures ranging from -15 to -75 °C).