The screw compressor can be visualized as a development of the gear pump. For gas pumping the rotor profiles are designed to give maximum swept volume and no clearance volume where the rotors mesh together. The pitch of the helix is such that the inlet and the outlet ports can be arranged at the ends instead of at the side. The solid portions of the screws slide over the gas ports to separate one stroke from the next so that no inlet or outlet valves are needed.
The more usual form has twin meshing rotors on parallel shafts (see Figures 4.19 and 4.23). As these turn inside the closely fitting casing, the space between two grooves comes opposite the inlet port, and gas enters. On further rotation, this pocket of gas becomes sealed from the inlet port and moved down the barrels. A meshing lobe of the male rotor then reduces the pocket volume compressing the gas, which is finally released at the opposite end, where the exhaust port is uncovered by the movement of the rotors. Various combinations of rotor sizes and number of lobes have been successfully employed. In most designs the female rotor is driven by the male rotor, and a study on optimizing rotor design for refrigeration applications is reported by Stosic et al. (2003). Maintenance of adequate lubrication is essential. Lubrication, cooling and sealing between the working parts is usually assisted by the injection of oil along the length of the barrels. This oil must be separated from the discharge gas and is then cooled and filtered before returning to the lubrication circuit (see Chapter 5).
Figure 4.19 Twin screw compressor rotors. The male rotor is on the left and the female on the right (Bitzer)
The single screw compressor has a single grooved rotor, with rotating star tooth seal vanes to confine the pockets of gas as they move along the rotor flutes (see Figure 4.20). Once again, various geometries are possible, but compressors currently being manufactured have a rotor with six flutes and stars with eleven teeth. The normal arrangement is two stars, one on either side of the rotor. Each rotor flute is thus used twice in each revolution of the main rotor, and the gas pressure loading on the rotor is balanced out, resulting in much lighter bearing loads than for the corresponding twin screw design. The stars are driven by the rotor, and because no torque is transmitted the lubrication requirements in the mesh are lighter also. Oil cooling and sealing is usual and the oil circuit is similar to that of the twin screw.
Arrows indicate suction gas flow
Figure 4.20 Single screw compressor (J&E Hall)
Screw compressors have no clearance volume, and there is no loss of VE due to re-expansion as in a piston machine. Volumetric losses result mainly from leakage of refrigerant back to the suction via in-built clearances. Oil is used for sealing, but leakage of oil, which contains dissolved refrigerant, reduces VE both by release of refrigerant and by heating the incoming gas. The VE decreases with increasing pressure ratio, but less than with some piston types (Figure 4.16). Leakage losses are a function of tip speed, so that smaller machines need to operate a higher speed to maintain efficiency. With synchronous motor drives, this sets a lower practical limit on the size (Figure 4.2).
In all screw compressors, the gas volume will have been reduced to a preset proportion of the inlet volume by the time the outlet port is uncovered, and this is termed the built-in volume ratio. At this point the gas within the screws is opened to condenser pressure, and gas will flow inwards or outwards through the discharge port if the pressures are not equal.
The absorbed power of the screw compressor will be at its optimum only when the working pressure ratio corresponds to the built-in volume ratio. The over and under compression losses can be visualized as additional areas on an indicator diagram as in Figure 4.21. This results in an IE characteristic having a strongly defined peak, as shown in Figure 4.18 . To the left of the peak, over-compression of the gas results in loss of efficiency, whilst to the right, there is under-compression with back flow of gas into the compression pocket
Figure 4.21 I ndicator diagram for compressors with built-in volume ratio, to illustrate over and under compression effects
When the discharge port is uncovered. Changing the size of the discharge port changes the position of the peak, and this is illustrated by the two curves in Figure 4.18. A screw compressor should be chosen to have a volume ratio suitable for the application. Leakage also contributes towards efficiency loss, but friction effects are quite small.
Capacity reduction of the screw compressor is effected by a sliding block covering part of the barrel wall, which permits gas to pass back to the suction, so varying the working stroke (Figure 4.22). It is usual for the sliding part of the barrel to adjust the size of the discharge port at the same time, so that the volume ratio is at least approximately maintained at part load. Many design variations and control methods exist. The single screw type will generally have two sliding valves; lifting valves are sometimes used instead of slides. Reduction down to 10% of maximum capacity is usual.
Figure 4.22 Capacity reduction slide for twin screw compressor (a) just starting to open,
(b) at minimum load. The valve is moved by a oil pressure acting on a piston, shown on the right (Howden)
The oil separation, cooling and filtering for a screw compressor add to the complexity of an otherwise simple machine. Liquid injection is sometimes used instead of an external oil cooler. Some commercial screw compressors have the oil-handling circuit built into the assembly. In Figure 4.23 the suction gas enters at the suction connection on the left, passes over the motor, through the compressor, into the multi-stage separator on the right and finally back to the discharge connection.
Figure 4.23 Semi-hermetic screw compressor with built-in oil separation (Bitzer)
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