Compressors temperature rise

The liquid-piston type of compressor has been of particular advantage when hazardous gases are being handled. Because of the gas-liquid contact and because of the much greater liquid specific heat, the gas-temperature rise is veiy small. 

Metallic Diaphragm Compressors (Fig. 10-99) These are available for small quantities [up to about 17 mvh (10 ftVmin)] for compression ratios as high as 10 I per stage. Temperature rise is not a... 

The gas turbine during the start-up is on an auxiliary drive, initially it is brought to a speed of about 1200-1500 RPM when ignition takes place and the turbine speed and temperature rise very rapidly. The bleed valves are open to prevent the compressor from surging. As the speed reaches about 2300-2500 rpm, the turbine is declutched from its start-up motor, the first set of bleed valves are closed, and then as the turbine has reached near full speed, the second set of bleed valves are closed. If the turbine is a two or three shaft turbine as is the case with aero-derivative turbines, the power turbine shaft will break loose at a speed of about 60% of the rated speed of the turbine. 

One significant practical difference in use of polytropic head is that the temperature rise in the equation is the actual temperature rise when there is no jacket cooling. The other practical uses of the equation will be covered as they apply to each compressor in the later chapters. 

The thermosyphon is a good system for remote areas where utilities are limited, but requires some careful design to ensure proper operation. This is a circulating system with the motive force derived from the change in densiiy of the cooling fluid from the hot to the cold sections of the sysieni. API dl8 pennits this system for discharge gas temperatures below 210 F or a temperature rise across the compressor of 150°F or less. 

The screw compressor can be evaluated using the adiabatic work equation. Discharge temperature can be calculated by taking the adiabatic temperature rise and dividing by the adiabatic efficiency then multiplying by the... 

The following equation provides a way of estimating the dischaige temperature if the shaft power is known, or it can be used to estimate the shaft power if the temperature rise and quantity of lubricant is known. The equation assumes 85% of the heat of the compressor is absorbed by the lubricant. 

The inverters are either voltage source or current source (see Figure 7-7a and b). There are other variations, but they apply to drivers smaller than the ones used with compressors. However, pulse-width-modulated (PWM) (see Figure 7-7c), transistorized units are less complicated and are relatively maintenance-free with reliable units available to at least 500 hp. For all but the smaller compressors, the current source inverter is the one typically used. With a six-step voltage source, a rule of thumb has been to size the motor at two-thirds of its rating so as not to exceed the insulation temperature rise. For current source motors, the output torque is not constant with decreased speed, which fortunately is compatible with most compressors, as torque tends to follow speed. For current source drives, one needs to upsize the motor captive transformer by approximately 15% to account for harmonic heating effects. 

Other sources, such as compression temperature rise, driver-induced vibration, or component problems (bows) can contribute to the machine shaking. These must be treated as they occur. As a minimum, care must be used to understand the nature of the sources to keep them from interacting with the resonant frequencies of the compressor. The best remedy to stop the excitation at the source. If this is not possible, selective tuning and proper application of damping must be used. 

Non-lubricated screw compressors have very close clearances and thus they are designed for limited ranges of discharge temperature, temperature rise, compression ratio, etc., all of which can cause changes in thesc clearances. Lubricated compressors have a somewhat broader tolerance to changes in operating conditions, but they are still more limited than reciprocating compressors. 

Polytropic Efficiency. This is the ratio of theoretical polytropic horsepower to actual brake horsepower at the compressor shaft. The polytropic efficiency does not include packing, bearing, or other losses. This efficiency is a measure of the hydraulic perfection of the compressor, and the value remains the same for any gas and for any speed (within reasonable limits). For an uncooled compressor, the polytropic, hydraulic, and temperature rise efficiencies are the same. ... 

As a general rule-of-thumh, the axial compressor will require about twice as many stages for a given requirement as the centrifugal compressor. The maximum number of axial stages is approximately 16. The temperature rise limitations as well as structural problems also limit the maximum stages for a given application. 

Heat generated by the compressor and prime mover must be dispersed. For air-cooled units sited in enclosed rooms this heat must be removed in order to limit the temperature rise. It is sometimes possible to recover this heat for use elsewhere. Intake openings should be located so that dust and other foreign matter do not enter with the air. 

The basic advantages of this process are (a) elimination of a mechanical device (recycle gas compressor) for controlling the adiabatic temperature rise, (b) combination of CO shift with methanation, (c) significant increase in byproduct steam recovery, and (d) significant capital advantages. 

Single-stage compressors can only be used for low pressure ratios. At high pressure ratios, the temperature rise will be too high for efficient operation. 

If we compare the work required to compress a given gas to a given compression ratio by isothermal and isentropic processes, we see that the isothermal work is always less than the isentropic work. That is, less energy would be required if compressors could be made to operate under isothermal conditions. However, in most cases a compressor operates under more nearly adiabatic conditions (isentropic, if frictionless) because of the relatively short residence time of the gas in the compressor, which allows very little time for heat generated by compression to be transferred away. The temperature rise during an isentropic compression is determined by eliminating p from Eqs. (8-17) and (8-19) ... 

The energy lost due to friction is actually dissipated into thermal energy, which raises the temperature of the gas. This temperature rise is in addition to that due to the isentropic compression, so that the total temperature rise across an adiabatic compressor stage is given by... 

The process in Figure 23-2 shows the compression of the ethylene in two stages. (There are more.) Ethylene will start to polymerize on its own in an uncontrolled fashion at 212°F, so in between compressors, the gas needs to be cooled. (Compression always makes the gas temperature rise. Thats why the bottom of your bicycle pump is hot after youVe filled your tire.)... 

The thermodynamic state of liquids is a changing polytropic one (close to isentropic) during quick compression (see Fig.4.1-1, for C02) and the corresponding temperature rise is low for liquids and large for gases (see AB in Fig. 4.1-1). The final compression temperature limits the compression ratio per stage with gas compressors. 

So far, we have limited our discussion to adiabatic compression efficiency. This sort of inefficiency downgrades work to heat. For a given compression ratio, the temperature rise of the gas as it flows through the compressor may be excessive, thus indicating a low adiabatic compression efficiency. Both centrifugal and reciprocating compressors suffer from this common problem, which is the subject of Chap. 30. 

A two-stage double-acting compressor with water cooled cylinder jackets and intercooler is shown in Figure 7.18(c). Selected dimensional and performance data are in Table 7.7. Drives may be with steam cylinders, turbines, gas engines or electrical motors. A specification form is included in Appendix B. Efficiency data are discussed in Section 7.6, Theory and Calculations of Gas Compression Temperature Rise, Compression Ratio, Volumetric Efficiency. 

First, it is instructive to examine the performance of a recuperated system that has only one compressor (i.e., remove the IC and C2 from Figure 8.2) and compare this to a simple cycle GT (i.e., also remove the recuperator from the diagram). Consider an isentropic compressor efficiency of 85%, isentropic turbine expander efficiency of 90%, recuperator effectiveness of 88% and no pressure losses. A fixed turbine inlet temperature of 1200 K will be assumed for various pressure ratios. This value is based on an assumed 1000 K SOFC inlet temperature, and a 200 K temperature rise from the SOFC inlet to the turbine inlet. The 200 K temperature increase from the cathode inlet to the turbine inlet is reasonable to assume given a cathode temperature difference across the cell of 150 K, and another 50 K temperature increase from anode exhaust combustion. Thus, 1200 K will be used as a base case for the turbine inlet temperature, and for sensitivity, values of 1100 and 1300 K will also be analyzed. 

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