Centrifugal compressors

Reciprocating compressors running at about 24 rev s_l have capacities of the order of 35 kW per cylinder when used for conventional air conditioning applications. The maximum number of cylinders on one machine is about 16, so this puts the maximum capacity of a piston machine at about 550 kW of refrigeration, excluding freak machines with large strokes and bores. Although one reciprocating compressor of this size is likely to be cheaper than a centrifugal machine of like capacity, the complication of controlling its output (cylinder unloading) may put it in an unfavourable light compared with a centrifugal

Centrifugal compressors

Suction sealoff a

Centrifugal compressors Centrifugal compressors

Centrifugal compressors

Centrifugal compressors

Centrifugal compressors Centrifugal compressors

Centrifugal compressors

Discharge 9 h

Fig. 12.12 Scroll compression process. Reprinted by permission of the American Society of Heating, Refrigeration and Air Conditioning Engineers, Atlanta, Georgia, from the 1992 ASHRAE Handbook—

Systems and Equipment SI.

Compressor, which can have a modulating control exercised over its capacity and chilled water flow temperature. Centrifugal systems are available for capacities as low as 280 kW of refrigeration, but they come into their own, economically speaking, at about 500 kW with a maximum at about 20 000 kW.

Whereas a reciprocating compressor is a positive displacement device, a centrifugal compressor is not. If the flow of gas into a piston machine is throttled it will continue to pump, even though the amount delivered is small, provided that its speed is kept up by an adequate input of power to its crankshaft. There is no ‘stalled’ condition. This is not so with a centrifugal compressor. The rotating impeller of a centrifugal compressor increases the pressure of the gas flowing through its channels by virtue of the centrifugal force
resulting from its angular velocity. The velocity of the impeller is constant in a radial direction but the linear velocity in a direction at right angles to the radius of the impeller increases as the radius gets greater. The energy input to the gas, which is rotated within the impeller, thus increases towards the periphery of the wheel. This input of energy is what causes the gas to flow outwards through the impeller against the pressure gradient, that is, from the low pressure prevailing at the inlet eye to the high pressure existing at the periphery. The function of the impeller casing, or the volute, is to convert the velocity pressure of the gas leaving the wheel to static pressure, with as much efficiency as possible.

In addition to the radial movement imposed on the gas by the impeller, the gas stream tends to rotate relative to the impeller. This is illustrated in Figure 12.13(a). On an absolute basis, any particular particle of gas will tend not to rotate but, since the wheel is rotating, the particle will appear to rotate relative to the wheel. The point Pj is facing the convex side of an impeller blade initially but, later in the process of rotation, denoted by P4, it is facing the concave side of the preceding impeller blade. The effect of this is to produce a circulatory movement of the gas within the impeller, as shown in Figure 12.13(b). It can be seen that this circulatory movement assists the flow to the periphery of the wheel, produced by centrifugal force, on the concave side of a blade but retards it on the convex side. The effect introduces losses which can be minimised by using a wheel with narrow channels between the impeller blades.

Centrifugal compressors

Fig. 12.13 Circulatory losses in the impeller of a centrifugal compressor.

For a given compressor, running at a given speed, the pressure-volume characteristic curve is virtually a straight line, as shown in Figure 12.14, if no losses occur. Losses do occur, however. They are the circulatory losses just described, losses due to friction, and losses caused by the fact that the gas entering the impeller has to change direction by 90 degrees, as well as having a rotation imposed on it. These entry losses may be modified by adjusting the swirl of the gas before it enters the inlet eye of the impeller. There is a proper angle of swirl for each rate of gas flow, that is, for each load. Variable-inlet guide vanes are fitted on all modern centrifugal compressors. Their position is adjusted to suit variations in the load, which permits a modulating control of output with little alteration in efficiency. The intention is that the machine should operate at a design point which involves the minimum loss, at the maximum efficiency.

A centrifugal impeller is designed to pump gas between the low suction pressure and the high condensing pressure. If the condensing pressure rises, the difference between these

Centrifugal compressors

Fig. 12.14 A pressure-volume diagram for a centrifugal compressor showing how the various losses

Give the characteristic curved shape.

Two pressures exceeds the design value and the compressor quite soon finds the task of pumping beyond its ability. Thus, whereas, the reciprocating machine will continue to pump, but at a steadily reducing rate as the condensing pressure rises, the pumping capacity of the centrifugal compressor falls rapidly away. This is illustrated in Figure 12.15(a). A similar behaviour is seen if the suction pressure is reduced, the condensing pressure being held at a constant value, as Figure 12.15(6) shows.

This feature of the performance of the centrifugal machine gives rise to a phenomenon termed ‘surging’. When the pressure difference exceeds the design pumping ability of the impeller, flow ceases and then reverses, because the high condensing pressure drives the gas backwards to the lower suction pressure. Pressure in the evaporator then builds up, and the difference between the high and low sides of the system diminishes until it is again within the ability of the impeller to pump. The flow of gas then resumes its normal direction, the pressure difference rises again, and the process repeats itself.

This oscillation of gas flow and rapid variation in pressure difference which produces it is surging. Apart from the alarming noise which surging produces, the stresses imposed on the bearings and other components of the impeller and driving motor may result in damage to them. Surging continuously is most undesirable, but some surge is quite likely to occur from time to time, unless a very careful watch is kept on the plant. This is particularly possible with plants which operate automatically and are left for long periods unattended. Surge is likely to occur under conditions of light load (when the suction pressure is low) coupled with a high condensing temperature.

The proper use of inlet guide vanes can give a modulating control of capacity down to 15 per cent or even, it is claimed, to 10 per cent of design full load.

The high heads necessary for air conditioning applications may be developed in two ways: either by running the impeller fast enough to give the high tip speed wanted or by using a multi-stage compressor. High tip speeds can be obtained by using large-diameter impellers, but if diameters are excessively large, structural and other difficulties arise.

Centrifugal compressors


Centrifugal compressors

Suction temperature (b)

Fig. 12.15 The performance of reciprocating and centrifugal machines may be shown in terms of

Both condensing and suction pressures.

High speeds of rotation are usually produced by using speed-increasing gears which multiply the 48 rev s’1 normally obtainable from a two-pole induction motor. Using an inverter to increase the frequency of the electrical supply is an alternative possibility. Full load running speeds lie in the range from 30 Hz to 1500 Hz with compression ratios from 2 to 30.

Although speed-increasing gear is noisy it has largely displaced two-stage compression, one reason being that gear ratios can be selected to suit the application, with maximum efficiency. Figure 12.16(a) shows a diagrammatic arrangement of two-stage compression with an ‘intercooler’ added. The purpose of this is to improve the performance of the cycle
by two expansion devices, A and B, and feeding some of the low pressure gas at state 9, obtained after valve A, to the suction side of the second stage compressor. Figure 12.16(6) describes the cycle in terms of the temperature-entropy changes involved.

Hot, high-pressure liquid leaves the condenser at a state denoted by 5. The pressure and temperature of the fluid is dropped by passing it through the first expansion device, A. Some liquid flashes to gas in the process, the flash gas being fed from the intercooler (merely a collection vessel) to the intermediate stage of compression. The remaining liquid passes from the intercooler through the second expansion device, B, and thence to the evaporator. Gas leaves the evaporator at state 1 and enters the first stage impeller, being compressed to state 2. When gas at state 2 mixes with gas at state 9, from the intercooler, it forms state 3 and enters the second stage impeller, leaving this at state 4. This then enters the condenser, and the process is repeated.

If the intercooler is not used, the cycle will follow the line 1 ^’-5-8-1. The refrigerating effect, represented by the area beneath 8′-l, is clearly less than the effect produced when an intercooler is used, represented by the area beneath 8-1.

Most condensers used with centrifugal machines are water-cooled but it is possible to get packaged air-cooled equipment in the range from about 450 kW to 1200 kW of refrigeration.

Posted in Engineering Fifth Edition