DOI: 10.1615/AtoZ.c.compressors

Compressors are devices for compression and delivery of gases. They are widely-used as separate units and as important parts of different types of heat engines. Compressors are driven by different types of prime movers (electric motors, steam and gas turbines, diesel engines, etc). There are two main principles of compressor action—the displacement principle with a cyclic process and the dynamic principle with a continuous process of gas compression. The pressure ratio (PR) for the displacement machine is the result of decreasing the volume that the trapped gas occupies. The dynamic machine develops pressure when power supplied to the rotor increases the angular momentum of the gas flowing through the rotor cascade. The stator cascade transforms this angular momentum of the gas into static pressure. There are also heat (thermosorptive) compressors in which pressure is increased after heating the matrix which absorbed the gas at lower temperature.

According to the developed pressure difference (PD), there are three types of compressors: low PD, medium PD and high PD units. If absolute pressure at the compressor inlet is significantly lower than atmospheric pressure, this unit is referred to a vacuum pump or turbomolecular pump. The inlet pressure may vary from 10−10 Pa to several Pa. Low PD compressors are termed blowers; the PR for them is equal to 1.05 to 1.15. For medium PD units, the PR is usually equal to 1.3 to 4.0. For high PD units, the PR is equal to 6 to 10 and more. Multicasing designs are used to achieve higher values of PR.

Displacement machines may be divided into reciprocating and rotary compressors and adsorptive units, while dynamic machines are divided into axial and radial flow compressors and jet units. Shear force radial compressors and magnetohydrodynamic pumps for plasma transportation by a moving electromagnetic field can also be classified as dynamic machines.

Reciprocating compressors

The reciprocating compressor, (Figure 1a), consists of a working cylinder (4) with two valves—a suction valve (3) and a discharge valve (2) in corresponding cavities in the cylinder head (1). The piston (5) moves reciprocatingly inside the cylinder with the help of the piston rod (6), crank (9), connecting rod (8) and cross-head (7). The discharge cavity is hermetically separated from the suction cavity. When the piston moves from the top, the cylinder is being filled with the gas at the suction pressure through valve 3. When the piston reaches its extreme position, valve 3 is closed. When the piston moves back, the cylinder volume is decreased, the gas pressure is increased, and the gas is compressed. When gas pressure exceeds the discharge pressure, valve 2 is opened and gas flows from the cylinder into the discharge line until the piston has reached the end of its stroke. This process takes place during one revolution of the shaft or double stroke of the piston. The extreme positions of the piston are called the top and bottom of the stroke. The volume between the top and the piston at its nearest point to the top position is a waste space. Reciprocating compressors can be one-stage and multistage. There can be single- and double-acting designs (in the latter, both sides of the piston are acting). Different positions of the cylinders are possible (vertical, horizontal, angular, V- or W-shaped, etc.).

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Reciprocating compressor.

Figure 1. Reciprocating compressor.

The working process of the ideal reciprocating compressor may be shown on a p − V diagram (Figure 1b). The suction process takes place at a constant pressure (line 4−1), then the gas is compressed inside the cylinder (line 1−2) and later is discharged (line 2−3). After valve 2 is closed, the gas pressure in the cylinder immediately drops from P3 = P2 to P4 = P1 and this process is repeated. Lines 4−1 and 2−3 are for constant gas parameters; only the volume and mass of the gas are changed. These lines are not the process lines. Line 1−2 is the compressing process line with a constant mass. The compressing process may be presented depending on the heat−exchange conditions. Line I is for the isothermic process (pV = const); line III is for the isoentropic process (pVγ = const); and line II is for the polytropic process (pVm = const). Polytropic compression work for the ideal compressor is calculated using the following equation:

For isothermal compression (m = 1), the work

where is the gas constant; , the universal gas constant; , the molar mass of the gas; and π = p2/p1, the pressure ratio. For isoentropic compression,

where γ is the specific ratio. In an actual reciprocating compressor, there are always some losses leading to the significant difference between the real diagram and the ideal one (Figure 1b). The existence of a heat flux between the cylinder wall and the gas should also be taken into account. The heat flux varies as far as its amount and direction are concerned. At the beginning of the compression process, heat is supplied to the gas (m > γ), while at the end of this process, the heat flux changes its direction (m < γ). This results in the increase of the indicator diagram area in p − V coordinates. The indicator diagram area A1 is proportional to the indicator compressor power Pi = Ain, where n is a rotative speed (RPM). The power absorbed by the compressor is equal to P = P1M, where the mechanical efficiency ηM = 0.82 to 0.95. The indicator power is related to the isoentropic and isothermic power by an expression PS,T = Pi · ηS−i,T−i, where ηS−i = 0.88 to 0.95 and ηT−i = 0.7 to 0.8.

The volume flow rate of the compressor is proportional to the cylinder volume Vc and n. It can be calculated as , where λν, is the coefficient of admission, (λν = 0.88 to 0.92). For atmospheric suction pressure, the reciprocating compressor can increase gas pressure to 10 MPa and more with a volume flow of from 10 to 200 m3/min. Reciprocating compressors are required for many gas storage, transmission, process, energy conversion and refrigeration systems.

Rotary compressors

There are many different types of rotary compressors. The principal types are the lobe (Roots type), screw (Lysholm type), sliding vane and gear (external and internal). Compared with reciprocating compressors, rotary compressors have better dynamic balancing, smaller sizes and mass, need no valves and have more uniform gas discharge.

The Roots type compressor has two rotors with parallel axes, (Figure 1a). Each rotor has two or three rounded lobes and revolve in opposite directions. A maximum clearance of 0.001D, where D is a rotor diameter, exists between the rotors and the casing. During rotation, the rotors transport the increments of gas from the suction side to the discharge side under an isochoric process. The volume flow rate for two lobe rotary compressor under suction condition is , where Ar, is the rotor cross-section area and l is the rotor length. The isochoric compression power is . The absorbed power P = PvM, where ηM = 0.85 to 0.95. The presence of radial clearances limits the allowable value of πc ≤ 1.8 for the rotary compressors without lubricant supply to the working cavity. Lubricant supply permits the increase of πc, but the lubricant has to be separated from the compressed gas.

Roots type compressor.

Figure 2. Roots type compressor.

The screw rotary compressor, Figure 2b, has different shape of rotor lobe and positions of the inlet and discharge holes. The compression process of the screw compressor is close to an isentropic one and takes place inside the compressor simultaneously with gas transport towards the discharge hole. The mechanical efficiency of the screw compressor is in the range ηM = 0.92 to 0.98 and the pressure ratio πc = 3 to 4. The tip velocity of the lobe on the leading rotor is as much as 120 m/s and more. Water injection inside the screw rotor reduces power consumption and leakage.

The process of compression in the sliding vane compressor is close to that of the screw compressor. This compressor has a rotor which revolves eccentrically inside a cylinder. The vanes (plates) can freely move inside radial or oblique slots. When the rotor revolves, these vanes come out of the slots and press on the internal surface of the cylindrical casing, forming cavities with changing volumes. The tip velocity of the vane is equal to 12 to 15 m/s. Gas compression, gas discharge and gas filling take place on successive arc lengths in the rotation.

The water-ring compressor, Figure 3, is similar in its process to the vane compressor but has some differences in design. A rotor with solid blades of the Rt tip radius revolves in the cylindrical casing with some eccentricity ε. Water or any liquid with low viscosity is in the space between the rotor and the casing. During the revolution of the rotor, the water assumes a ring shape with the internal radius R. The minimum depth of the blade immersion is equal to a. The point B indicates a minimum distance between the hub of the rotor and the water-ring. Gas is sucked through the large port in the cylinder flange (shadowed) and is discharged through the smaller port. Working cavities are located between the rotor hub and the water-ring. The compression process is close to the isothermic one. These machines have long service life and are suitable for any dusty gas compression. They are widely-used as vacuum-pumps and create up to 98% vacuum. Their pressure ratio may be up to 2.5 to 2.8 when they are used as blowers; but the efficiency is only 40 to 45% due to high losses because of the interaction between the blades and the water-ring and due to friction in the casing.

Figure 3. Water ring compressor.

The thermosorptive compressor uses the ability of some substances or metal alloys to absorb some gases at room temperature and pressure, with subsequent discharge of the gases at much higher pressure after some slight heating of the matrix. For instance, the LaNi5 alloy is capable of absorbing hydrogen-forming LaNi5x hydride. After the slight heating of the powdery matrix, the hydrogen leaves at an elevated pressure. A similar absorptive process can take place when methane is dissolved in water to form a hydrate. A slight heating of the hydrate releases the methane at higher pressure.

The axial flow compressor, Figure 4, consists of a casing (1) with the inlet (8) and outlet (10) ducts, supplied with the axisymmetric inlet confuser channel (7) to obtain uniform flow (2) and an axisymmetric outlet diffuser channel (9) for partial transformation of the kinetic energy of the gas flow into static pressure. The rotor has two bearings (12) and sealings (11) to reduce the leakage of the gas between the rotor and the casing. Compressor blading usually consists of several stages placed one after another. A stage has two rows of the wing-shaped blades, which form aerodynamic cascades for the rotor (R) (3) and the stator (S) (4). Both cascades have curvilinear diffuser channels. In front of the first rotor blade row, there is often an inlet guide vane (IGV) row with the turbine-type cascade. At the entrance to the outlet axisymmetric channel, there is an additional stator (S') blade row (6) to get the axial flow in front of the outlet duct.

Axial flow compressor.

Figure 4. Axial flow compressor.

Figure 5a shows the distribution of the static pressure p, the absolute velocities v and the relative velocities w of the gas for several cascades along the compressor axis. The axial velocity u is related to w and v via the vector triangle (see Fig. 5b). From the continuity equation, one can see that the increase of the gas density ρ leads to a decrease of the blade height towards the compressor exit. The main parameters of the cascade are solidity σ = c/s [ratio of the blade chord c to the blade spacing (pitch) s], the stagger angle ξ and the camber angle θ for the profile mean line. Compression work of the stage can be calculated either by the expression

where Cp is the specific heat at constant pressure πst, the stage pressure ratio and πst, the stage efficiency. The approach of relative velocity to the speed of sound leads to decrease of the stage efficiency η . The pressure ratio for the subsonic stages is less than 1.2 to 1.25 and for the transonic stages, about 1.75 to 2.0 for the efficiencies of 0.88 to 0.92. Generallyspeaking, the stage efficiency is influenced by the flow coefficient (the ratio of the mean flow rate velocity to the tip velocity Ut of the blade); the head coefficient (the ratio of the work of the stage to ); and the degree of reaction (the ratio of the rotor static pressure increase to the static pressure increase). The vortex and the constant degree of reaction designs are widely-used for axial flow compressors. The multistage axial flow compressors can be designed to achieve the total pressure ratio in one casing of the order of 10 to 15 and efficiencies of 0.86 to 0.88.

Figure 5. 

The radial compressor consists of a casing of a complicated shape, in which the rotor induces a radial flow towards an outlet stator ring.

Due to the action of centrifugal forces a radial stage produces a higher pressure ratio than the axial one, in which the compression is mainly connected with the diffusion effect at the cascade. The efficiency of the radial compressor stage is usually equal to 0.79 to 0.83 for a pressure ratio of 3 to 5.

Static pressure in the flow can be increased with the help of a jet device which acts as the injector.

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