This latter characteristic can be better explained by reviewing test data or performance maps of typical centrifugal and positive displacement-type machines. Figure 1 depicts performance mapping of an available positive displacement-type machine, in this instance, a rotary screw compressor. Figure 2 shows performance mapping of a centrifugal-type compressor.
Although multivariate data are displayed in these maps – and they may appear quite complicated at first blush – the mapped performance differs substantially between these two technologies.
For the screw compressor, constant speed operating lines appear almost vertical. Very small changes in flow rate result in large changes in pressure while operating at constant speed. Maintaining appropriate discharge pressure and, hence, nozzle velocity relies on precisely defined downstream flow geometry.
As a result, variable speed control becomes necessary to “trim” operation to the precisely correct discharge pressure.
For the centrifugal machine, the characteristic is the exact opposite. (See Figure 2.) Constant speed operating lines are seen as nearly horizontal. Very large changes in flow have little affect on discharge pressure. So a single compressor can accommodate a wide range of flow with relative stability of discharge pressure while operating at constant speed. Impeller speed is the knob for controlling discharge pressure alone, and there is little concern over precisely matching the downstream flow geometry. Also, a system can be enhanced to attain higher flow operation as long as sufficient power is available without the need to change the compressor component.
Other differences can be seen when comparing overall package size and weight and power consumption. The latter is directly influenced by compression efficiency, a by-product of compressor design and technology choice. The performance maps of Figures 1 and 2 depict efficiency performance of the screw and centrifugal units respectively. Although these data are peculiar to specific products, they may be taken as representative for our purposes.
Clearly, the centrifugal unit exhibits generally higher operating efficiency – 79 percent peak vs. 65 percent peak for the screw product. But does a 15-point efficiency difference really matter? What impact does this have on overall system requirements and operating costs? Moreover, at what efficiency is the compressor operating at the particular flow and pressure point of interest?
Fortunately, efficiency differences and their overall affect on power and, hence, operating costs can be readily evaluated via the polytropic compression equations. Table 1 summarizes such a study with performance taken at the 1.9 PR and 100 PPM air flow operating conditions. For these calculations, inlet air is taken at 20 degrees F and 14.50 psia, typical of conditions during deicing operations.
Contrasted are three compressor technologies including the Figure 1 and Figure 2 screw and centrifugal devices, and a third representing a rotary lobe roots device typically found in a wide variety of industrial applications. Also, even though a given device or technology may perform at a certain peak efficiency level, what is more important is the efficiency at the specific pressure and air flow operating point of interest.
For example, the Figure 1 twin-screw device attains a peak 65 percent efficiency, but is only 54 percent efficient at the 1.9 PR – 100 PPM flow point of interest. The Figure 2 compressor is closer – 76 percent operating efficiency against a best of 79 percent.
This reveals the importance of close scrutiny of mapped performance of any given compressor technology or device, and whether it is truly optimum for the specific air supply needs. This raises the challenge to develop compressor stages that are actually tuned so that peak efficiency performance is attained specifically at the desired pressure and flow operating condition. This assures the absolute minimum power and, hence, the lowest operating cost condition for the forced air deicing equipment.
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