It is sometimes hard to imagine that the piping plays any role in the operation of a refrigerating plant, as it is totally static and has no moving parts. However, that does not mean there is no motion connected with it. The motion that does occur is the flow of the refrigerant within. As a result of that flow, friction causes there to be pressure changes within the pipes: the greater the flow, the greater the pressure necessary to cause the flow. The effect on the operation of an actual system is profound, particularly on that of a large system.
Practitioners who work with systems regularly refer to system operating conditions in a way that ignores the piping pressure differences. For example, the system condensing pressure is referred to as both the pressure that exists in the condenser as well as that at the compressor discharge. Converting the pressure to saturation temperature, the condensing pressure will be referred to as 95°F. The compressor will likewise be considered to be operating at a saturated discharge temperature of 95°F. Of course, the fact is that both parts of the system cannot be at the same temperature. Because the refrigerant gas must flow through the pipes to get from the compressor to the condenser, the pressure at the compressor must be greater than that in the condenser.
Similarly, air-cooling units operating in higher-temperature refrigerated spaces will be said to be operating at a saturated evaporating temperature of, for example, 20°F. Commonly, the compressors will be referred to as operating at that same saturated suction temperature (SST). But, they cannot. There must inherently be a lower pressure at the compressor than at the evaporator(s), and depending on the size of the plant and the type of system that the refrigerating plant is, the pressure at the compressor can be considerably lower than that at the evaporator(s).
The reason this is important is that the lower the saturated suction temperature and the higher the discharge (saturation) temperature (SDT) at which a compressor operates, the lower is its capacity and the poorer is its efficiency. For example, a 400 cfm displacement reciprocating compressor, operating at 20°F SST and 90°F SDT, has a capacity, with ammonia as the refrigerant, of 131.6 tons of refrigeration (TR). Its specific power consumption is 1.046 horsepower per ton (hp/T).
At 15°F SST and 90°F SDT, the capacity of the same compressor is 115.7 TR and the specific power consumption is 1.154 hp/T. At 10°F SST and 90°F SDT, the capacity is 100.5 TR and the specific power consumption is 1.277 hp/T. In an ammonia system, at constant SDT, the loss in compressor capacity is at a rate of approximately 2.5 percent for each 1°F reduction in saturated suction temperature. If it is decided that the cooling units are to operate at an evaporating temperature of 20°F, what determines what the actual compressor SST will be? It is the pressure loss in the piping (∆P) between the evaporators and the compressor, and that is the designer's choice.
The pressure differences at the low-temperature levels of freezing plants can have an even more dramatic effect on the performance of compressors. For example, at 20°F, the vapor pressure/temperature relationship for ammonia is 1°F per psi change in pressure. A common operating temperature for process freezers is -40°F. At -40°F, the vapor pressure/temperature relationship for ammonia is 3.4°F/psi. Thus, a 1 psi ∆P at 20°F will affect the compressor capacity by approximately 2.5 percent. However, a 1 psi ∆P at -40°F will affect the compressor capacity by 8.5 percent.
A similar analysis on the discharge side of compressors will show that, at a constant SST, the ∆P in the discharge piping - that is, the piping between the compressor(s) and the condenser(s) - affects the compressor by an approximate 0.5 percent reduction in capacity for each 1°F increase in SDT. There is also an associated reduction in efficiency. The specific power consumption increases by 1.2 percent for each 1°F increase in SDT.
Intuitively, it would appear that the objective would be to make the pipes as large as possible to minimize the adverse effects of reduction in compressor capacity -- or, conversely, having to use larger, more costly compressors -- and lower efficiency. Richards pointed out, however, that for most refrigerant piping, there is an economically optimum size of pipe. IIAR provides discussion of how to analyze the piping needs for any system and how to size the piping according to the economic principles proposed by Richards.
It is important to understand how the various elements of a piping system contribute to the overall pressure loss when a refrigerant flows through it, and most important, how compressors are rated and at exactly what temperatures a manufacturer's capacity rating applies. Cole presented a discussion of compressor ratings relative to system piping pressure losses. PCE
Headquartered in Arlington, Va., the International Institute of Ammonia Refrigeration is an international association serving those who use ammonia refrigeration technology through education and advocacy. Its membership includes end users, contractors, engineers, equipment manufacturers and others in the industry. Contact IIAR at (703) 312-4200 or www.iiar.org.
ReferencesWilliam V. Richards, “Refrigerant Line Sizing not Dependent on Length,” Proceedings of Commission B2, 16th International Congress of Refrigeration, IIR, p. 240-244, 1983.
Ammonia Refrigeration Piping Handbook, International Institute of Ammonia Refrigeration, Washington, D.C., 2000.
Ronald A. Cole, P.E., and Jason R. Cole, “Compressor Ratings: What Do They Mean?” Heating/Piping/Air-Conditioning, p. 41-42, 49-50, February 1993.