In a centrifugal compressor, energy is transferred from a set of rotating impeller blades to the gas. The designation “centrifugal” implies that the gas flow is radial, and the energy transfer is caused from a change in the centrifugal forces acting on the gas. Centrifugal compressors deliver high flow capacity per unit of installed space and weight, have good reliability, and require significantly less maintenance than reciprocating compressors. However, the performance characteristic of centrifugal compressors is more easily affected by changes in gas conditions than is the performance of reciprocating compressors. On this page, the performance characteristic curve is presented with emphasis on process control of capacity by speed variation, suction throttling, or variable inlet guide vanes. Process control to avoid operation in a damaging surge condition is also addressed.
The physical size (diameter) of a centrifugal compressor is determined by the volumetric flow rate at the inlet. The compression ratio (or head) determines the number of stages (length). The rotating speed of a centrifugal compressor is an inverse function of diameter to maintain a desired peripheral speed at the outer diameters of the impellers regardless of the physical size of the compressor. Very large (i.e., high-volume) flow compressors may operate at speeds as low as 3,000 rpm. Conversely, low-volume flow compressors may operate at speeds up to 30,000 rpm. Power requirement is related to mass flow, head, and efficiency. Depending on the particular application, centrifugal compressor powers can range from as low as 500 hp (400 kW) to more than 50,000 hp (40 MW).
At low volume flow rates, the width of the gas passages in a centrifugal compressor becomes narrow, and the effects of friction become significant, resulting in reduced efficiency. For this reason reciprocating compressors often are more appropriate for low-volume flow applications. For further discussion of this subject, see the section below on compressor selection.
The API has produced an industry standard, API Standard 617, which is frequently used to govern the design and manufacture of centrifugal compressors. A typical centrifugal compressor package is shown in Fig. 1. The compressor shown is mounted on a single baseplate and is driven by an electric motor.
Multi and single stage centrifugal compressors
Multi-stage centrifugal compressors
Multistage centrifugal compressors can be arranged in a variety of flowpath configurations employing from one to ten impellers, depending on the head required for the process duty. When intercooling is not needed, the arrangement is usually a straight-through (inline) configuration. For applications that require intercooling, the resulting two-section compressor may be configured in either an inline (compound) or back-to-back arrangement. For high-flow/low-head applications, a double-flow configuration is sometimes employed. In a double-flow arrangement, half of the flow enters the compressor through an inlet connection at each end of the casing and exits the casing through a common discharge connection in the center. All of these configurations described are beam-type designs in which the impellers are located between the radial bearings.
Single-stage centrifugal compressors
Single-stage centrifugal compressors may be configured as a beam design or with an overhung impeller arrangement. In the overhung configuration, the impeller is located at the nondrive end of the shaft (outboard of the nondrive end radial bearing).
The major components of various centrifugal compressor flowpath configurations are illustrated in Fig. 2 through Fig. 5. This section describes the major elements of centrifugal compressors.
Case (casing or housing)
The case (casing or housing) is the pressure-containing component of the compressor. The case houses the stationary internal components and the compressor rotor. Bearings are attached to the case to provide both radial and axial support of the rotor. The case also contains nozzles with inlet and discharge flange connections to introduce flow into and extract flow from the compressor. The flange connections must be properly sized to limit the gas velocity as necessary. The case is manufactured in one of two basic types:
- Vertically split
Construction can be cast (iron or steel), forged, or fabricated by welding.
Horizontally (axially) split case
A horizontally split case is split parallel to the axis of the rotor. The upper half of the case is bolted and doweled to the lower half. Access to the internals of the compressor for inspection and maintenance is facilitated with this case design (especially when the process piping connections are located on the bottom half of the case). The horizontally split design is inherently pressure-limited to prevent gas leakage at the case split joint.
Vertically (radially) split case
This case is split perpendicular to the axis of the rotor. Heads (end covers) are installed at both ends for pressure containment. The vertically split case configuration is capable of handling higher pressures than the horizontally split type. The rotor and stationary internals are assembled as a cylindrical inner bundle that is inserted axially through one end of the case. Inspection and maintenance of a radially split centrifugal compressor require that the inner bundle be removed for disassembly. Removal of the inner bundle requires that sufficient space be provided in the layout of the compressor installation.
The compressor rotor is fundamentally an assembly of impellers mounted on a steel shaft. Additional rotor components include miscellaneous hardware, such as:
- A thrust balance drum (balance piston)
- Impeller spacers
- Seal sleeves
- A thrust disc
- One or two couplings
A typical compressor rotor is pictured in Fig. 6.
The impellers impart velocity to the gas with blades that are attached to a rotating disc. The impeller blades are forward-leaning, radial, or backward-leaning (with respect to the direction of rotation) depending on the desired performance characteristic curve. Backward-leaning blades tend to provide the widest operating range with good efficiency. They are the most commonly used blade shape. Proper sizing of the impeller flow channels is determined by the volumetric flow rate to control gas velocities through the impeller. This means that, in a multistage compressor, the impellers must be properly sized for peak performance and properly matched to accommodate the volumetric flow rate reduction through the compressor. Impellers can be of the open type without a cover plate or the closed type that incorporates a cover plate attached to the blades. Most multistage compressors use the closed-type impeller design. Impeller construction can be:
- Electron beam welded
- Welded conventionally
For most applications, high-strength alloy steel is selected for the impeller material. Stainless steel is often the material of choice for use in corrosive environments. Because the impellers rotate at high speeds, centrifugal stresses are an important design consideration, and high-strength steels are required for the impeller material. For gases containing hydrogen sulfide, it is necessary to limit the impeller material’s hardness (and therefore strength) to resist stress corrosion.
Multistage centrifugal compressor rotors have natural resonant frequencies that must be outside the operating speed range. Rotordynamic design considerations can limit the maximum number of stages per case or, stated another way, limit the maximum speed for a given number of stages.
After the gas enters the compressor through the inlet nozzle, it must be directed to the inlet of the first stage impeller in a way that uniformly distributes the flow to the impeller at a desired velocity. A system of internal stationary components is designed to deliver the gas to the first impeller with minimal pressure drop. Stationary inlet guide vanes are normally positioned adjacent to the impeller inlet. Variation of the inlet guide vane angles can be employed to adjust the flow capacity of the compressor’s performance characteristic curve. However, a variable inlet guide vane system introduces mechanical complexity as well as additional sealing considerations (see the section below on Flow Control)
The gas exits the impeller at high velocity and enters a diffuser passage. The diffuser is an important part of the stationary flowpath that usually comprises two brllel walls forming a radial flow channel. In the diffuser, the gas velocity decreases and dynamic pressure is converted to static pressure. Diffusers can be either vaneless or vaned. After exiting the diffuser passage, the flow encounters a return bend, which creates a 180-degree turn in the direction of flow (i.e., from radially outward to radially inward). Following the return bend, the flow enters a vaned return channel that directs the flow inward to the next impeller. The function of the return channel is (in the same manner as the first stage inlet system) to uniformly deliver flow to each impeller with minimal losses. The inlet guide vanes are located at the exit of the return channel. The components that form the return channel are called “diaphragms,” and the diffuser passages are the spaces between adjacent diaphragms. Inlet guide vanes can be attached to a separate piece fitted into the diaphragm or an integral part of the diaphragm.
Following the last stage impeller, the gas must be collected and delivered to the discharge flange. The stationary component, typically used for this purpose, is a discharge volute. The volute must be well matched to the discharge nozzle to minimize pressure losses. Discharge nozzle velocities also must be kept within limits to avoid excessive noise levels. All of the stationary components previously described play an important role in overall compressor performance.
Bearings and seals
Centrifugal compressors are equipped with two radial (journal) bearings to support the rotor weight and position the rotor concentrically within the stationary elements of the compressor. One thrust bearing also is used to ensure that the compressor rotor is maintained in its desired axial position. The thrust bearing usually is a “double-acting,” tilt-pad design installed at both sides of a rotating thrust disc. Proper rotor axial position is thereby assured regardless of the direction of the net axial pressure forces acting on the rotor.
Two distinct categories of compressor seals are used:
- Internal seals
- Shaft seals
Internal seals minimize internal recirculation losses between stages and across the thrust balance drum. Labyrinth type seals are customarily used for this purpose to maximize operating efficiency.
Shaft seals are required to seal the gas inside the compressor at the point where the compressor rotor shaft penetrates the case. This vital sealing function is necessary to prevent escape of process gas to the environment surrounding the compressor. Dry gas seals are the most commonly used type of shaft seal. Liquid film seals are sometimes used.
Labyrinth-type seals are used to minimize recirculation losses within the compressor. A labyrinth seal consists of a number of teeth (knife-edges) that can be either stationary or rotating. Stationary labyrinth teeth are fitted to the compressor stationary components very close to the compressor rotor (see Fig. 7). Sealing action is the result of flow resistance caused by repeated throttling across the labyrinth teeth. Labyrinth seals are designed so that one of the two adjacent parts (labyrinth teeth and rotor) is relatively soft. The softer material yields on contact without damage to the harder material. Compressor manufacturers select labyrinth seal clearances that are as tight as practical to minimize leakage while avoiding heavy rubbing with the rotor.
Dry gas seals
Beginning in the late 1980s, the compressor industry began to embrace the application of dry gas seal technology to the critical function of shaft sealing. The seal consists of a rotating disc running very close to a stationary ring. The rotating disc face contains special grooves that generate an axial (“lift”) force during rotation. The stationary ring is backed by a quantity of coil springs that force it tightly against the rotating disc when the compressor is at rest. The lift force compresses the coil springs slightly, resulting in the very small running clearance between the two faces. This small clearance effectively limits gas leakage from the compressor seals. The small amount of gas leakage exits the compressor through auxiliary seal piping, where it is then either sent to a flare system or to some other recovery system. Usually the two compressor seals (inlet and discharge ends of the compressor) are subjected to the gas suction pressure. A thrust balance line (see further discussion in the section on bearings below) subjects the discharge end dry gas seal to inlet pressure, thereby avoiding the need to seal the higher discharge pressure.
Dry gas seals require clean and dry gas for reliable operation. Seal gas is normally taken from the compressor discharge and then cooled and filtered as part of an external seal gas processing system. A seal reference pressure is measured just inboard of the dry gas seal, and a pressure regulating valve supplies the seal gas to the sealing faces at a pressure slightly above the reference pressure. This system ensures that the seals are not exposed to untreated process gas containing liquids or particulate matter that could damage the seals. Although dry gas seals are relatively expensive, their auxiliary system is less complex, physically smaller, and less expensive than the auxiliary system required by the predecessor liquid film design.
Liquified film seals
Liquid film seals can be of the bushing type or mechanical contact type. The bushing type is a very simple and rugged design that incorporates two adjacent seal rings (bushings) at each end of the compressor. A sealing fluid is introduced into the space between the seal rings at a pressure slightly above the process gas pressure inboard from the inner ring. The pressure differential across the inner ring is assured by an overhead seal oil tank pressurized by compressor suction pressure. The elevation above the compressor of the oil level in the tank assures the required seal ring pressure differential. For almost all centrifugal compressors equipped with liquid film seals, the sealing fluid is the same light turbine oil as that used to lubricate the bearings. Therefore, the auxiliary seal oil system needed to supply the seal oil can be combined with (or separate from) the auxiliary lube oil system.
The inner seal ring is designed to minimize oil leakage into the process side. Inner seal leakage (also called sour oil leakage) mixes with the process gas and is drained from the compressor as an oil/gas mixture. A labyrinth seal inboard of the sour oil drain is installed to prevent seal oil from contaminating the process gas. The oil/gas mixture drains into a degassing tank where the gas is removed so that the oil can be sent to a seal oil reservoir for reuse.
The outer seal ring breaks serve to inhibit flow as pressure is reduced to an atmospheric drain. This drain is common with the bearing oil drain when a combined oil system is used. When the lube and seal oil systems are separate, a buffered labyrinth seal is placed between the lube and seal oil drains to ensure that there is no oil carryover from one system to the other.
Mechanical contact seals employ a stationary carbon ring against a rotating seal face. Oil is also used as the sealing medium in mechanical contact seals. The sealing oil is introduced by a pressure-regulating valve that is maintained at 25 to 40 psi above the seal reference pressure. One advantage of mechanical contact seals is a significantly reduced sour oil leakage compared with the bushing design. Unlike oil film seals, mechanical contact seals can be supplied with a feature that allows the compressor to maintain case pressure during shutdown without requiring that the auxiliary seal oil system be operating. Mechanical contact seals, however, are relatively complex.
The radial bearings most often used in centrifugal compressors are the tilting pad type and are continuously lubricated with light turbine oil. Before tilting pad designs, sleeve type bearings were commonly used. The tilting pad bearing design provides rotor-dynamic characteristics that help assure smooth and reliable mechanical operation. Radial bearings are sized to be large enough to support the rotor weight, yet small enough to operate at sufficiently low peripheral speeds required to limit operating temperature to acceptable levels. Some centrifugal compressors are equipped with magnetic radial bearings. These bearings suspend the rotor by electromagnetic force to center the rotor within an air gap at the bearing. Use of magnetic bearings eliminates the need for an auxiliary lube oil system; however, the magnetic bearing control system also requires cooling.
The pressure rise in each of the stages of a centrifugal compressor creates an axial thrust force that acts toward the inlet end of the compressor. Depending on the overall pressure rise in the compressor, these thrust forces can be significant. An inline configuration employs a thrust balance drum (balance piston) to generate a thrust force to oppose (“balance”) the sum of the impeller thrust forces. Located at the discharge end of the compressor, the balance piston is a simple disc-shaped element installed on the compressor shaft and equipped with a seal around its outer diameter. The space adjacent to the outboard face of the balance piston is subjected to compressor suction pressure as created by an auxiliary thrust balance line. The inboard surface of the balance piston is subjected to what is essentially the compressor discharge pressure. The resulting pressure differential across the balance piston creates an axial force toward the discharge end, thus opposing the impeller thrust forces. Proper selection of the balance piston diameter results in small net thrust force and allows use of a reasonably small thrust bearing to absorb the residual thrust forces and maintain proper rotor axial positioning.
Like the radial bearing, the thrust bearing is usually a tilting-pad design lubricated with light turbine oil. Some thrust bearing designs employ a system of leveling blocks behind each tilting pad to ensure uniform load distribution. As with the radial bearings, magnetic thrust bearings also are available.
The performance characteristic of a centrifugal compressor is graphically presented in the form of a family of curves that collectively are known as a performance map or operating envelope. An example of a performance map is given in Fig. 8. In the example map, the inlet volume flow is plotted along the x -axis and the head or pressure ratio is plotted along the y -axis. The “approximate surge limit” depicted at the left side of the map defines the minimum flow necessary to avoid a potentially damaging surge condition (see section on Surge below). At the extreme right portion of the map is the “stonewall” (choke) limit (see section on Stonewall (Choke) below). Each of the family of curves from the surge limit to the stonewall represents the flow vs. pressure characteristic at a given compressor speed. The slope of the curve varies with the number of stages, becoming steeper with an increasing number of stages. The elliptical curves (dashed lines) denote compressor efficiency. The design point is at 100% speed, and the compressor components are selected so that the design point has a safe margin from surge and stonewall, as well as optimum efficiency.
The surge limit defines the flow at which, for a given speed, the operation of the compressor becomes unstable. At flow rates below the surge limit the characteristic curve actually droops toward zero flow after having reached its maximum point at the surge limit. Because operation below the surge limit is unstable, this portion of the curve is not shown in Fig. 8. When the flow is reduced below the surge limit, the pressure at the discharge of the compressor exceeds the pressure-making capability of the compressor, causing a momentary reversal of flow. When this flow reversal occurs, the pressure of the discharge system is reduced, allowing the compressor to resume delivering flow until the discharge pressure again increases, and the surge cycle repeats. Surging usually creates a clearly audible noise. Prolonged operation in this unstable mode can cause serious mechanical damage to the compressor. When operating in a surge condition, the compressor discharge temperature increases significantly and the compressor experiences erratic and severe vibration levels that can cause mechanical damage—particularly to the internal seals.
A compressor can be brought out of surge in a number of ways. The most obvious is to increase flow (see section on Antisurge Valves below). Decreasing discharge pressure and/or increasing speed are other ways to move out of a surge condition.
Compressor manufacturers usually perform an aerodynamic performance test before delivering the compressor. Determination of the compressor’s actual surge limit is a very important aspect of the manufacturer’s shop testing program.
The stonewall limit of the performance curve defines the flow at which the gas velocity at one of the impellers approaches the velocity of sound for the gas at the conditions within the compressor where this sonic condition is first encountered. At the stonewall (or choke) flow the pressure vs. volume curve becomes essentially vertical, and it is not possible to develop head or pressure at any greater flow. When the required operating flow exceeds the stonewall limit, the only remedy is to reconfigure the compressor with impellers (and matched stationary hardware) designed for larger flow rates.
A centrifugal compressor may be configured with one of a variety of process connection arrangements. For grade-mounted installations, the process connections are most often positioned on the upper half of the casing with the process piping connected from above the compressor. In some installations, horizontal (side) connections are employed. The horizontal connection arrangement is frequently used in booster compressors for gas transmission. A radial split case design (see the section on Case above) is preferred for these two arrangements. Another arrangement is the mezzanine-mounted configuration. With this type of installation, the compressor connections are on the lower half of the casing, and the process piping is connected from underneath the compressor. If the operating pressures are sufficiently low, an axial split case design is appropriate (see the subsection on cases). To achieve optimum performance, it is necessary to install the compressor with a sufficiently long straight section of inlet piping upstream of the compressor inlet flange. Most compressor manufacturers require that the length of this straight section be at least two times the inlet flange diameter.
The compressor must be well integrated into the entire process so that start-up, operation, and shutdown can be safely controlled. The next section provides a description of control concepts and the process equipment required. (Refer to Fig. 9 for a typical process flow diagram.) The individual elements of the control and safety system illustrated in Fig. 8 are discussed in this section.
Most compression processes require the compressor to deliver a relatively constant discharge pressure over a range of capacities. However, the centrifugal compressor characteristic curve from Fig. 8 shows that the pressure ratio, in fact, varies continuously with flow. The process can control either suction or discharge pressure. If one is fixed, the other will vary as dictated by the compressor characteristic curve. The three methods of maintaining a constant discharge pressure for varying capacity are discussed next.
Centrifugal compressor drivers are either of the fixed or variable speed type. Most steam or gas turbines and those electric motors equipped with a variable frequency drive system are all available as variable speed drivers. For a given discharge pressure, compressor capacity may be increased by merely increasing the speed of rotation. Conversely, capacity may be decreased by reducing compressor speed. Capacity control by speed variation is the most effective way to maximize the operating flexibility of a centrifugal compressor.
Suction throttle valves
A fixed-speed motor is often the least expensive driver for a centrifugal compressor. When designing a centrifugal compressor driven by a fixed-speed motor, it is necessary to establish the speed based on the operating condition that requires the largest capacity for the required discharge pressure. When operating at lower capacities, the compressor inherently delivers a greater discharge pressure (for a given process suction pressure) than desired. The solution to this problem is to install a throttle valve at the inlet of the compressor. Suction pressure reduction by throttling increases the pressure ratio required to deliver a given discharge pressure. The economic trade-off for this method of capacity control is additional compressor power vs. additional capital expenditure for a variable speed driver.
Variable inlet guide vanes
As discussed in the section on Stationary Components above, the compressor performance characteristic curve can be adjusted by changing the direction of the flow of gas into the impeller. When a system of variable inlet guide vanes is employed, it is possible to adjust the inlet guide vane angles to maintain a desired discharge pressure over a range of capacity. Practical design limitations make it difficult to install variable vanes at all stages other than the first stage. For single stage compressors, this method of control is sometimes quite effective. However, for multistage compressors, the range of control is less effective and becomes even less so with increasing numbers of stages.
As discussed in the section on Surge above, avoiding surge is extremely important. The installation of an antisurge (recycle) valve and its associated control devices is required. The antisurge valve is located in a recycle line connecting the compressor discharge to the inlet. For multisection compressors, it is good practice to install a separate recycle line with an antisurge valve for each of the compressor sections. Instrumentation is required to measure the flow to each section, and a surge controller must initiate the opening of the recycle valve when reduced capacity approaches the surge limit. The capacity at which the antisurge valve begins to open is usually set to be about 10% larger than the actual surge limit.
For variable speed compressors, the surge limit curve (see Fig. 8) defines the relationship between the surge limit and the operating speed. The logic programmed into the antisurge controller maintains the 10% safety margin, regardless of speed. This can be depicted graphically by a line parallel to the surge limit curve and is typically called the “control line.”
The gas recycled through the antisurge valve also must be cooled because its source is the compressor discharge. If uncooled, the suction temperature will increase by mixing the hotter recycled gas with the main process inlet gas.
The flare valve protects upstream equipment from overpressurization that may occur because of a flow increase and prevents overloading of the compressor driver. For a constant discharge pressure system, an increase in flow results in an increase in suction pressure. Higher suction pressures deliver more mass flow and, therefore, increase the power required to operate the compressor. The presence of a suction throttle valve also can contribute to an increase in pressure upstream from the compressor. Thus, flare valves are particularly important in installations with inlet throttling.
Shutdown valves are installed at both the suction and discharge to enable the compressor to be isolated during shutdown periods. To satisfy safety concerns, the shutdown valves should be located outside any building or enclosure. Automatic control of the shutdown valves is usually employed.
At shutdown, after the shutdown valves have isolated the compressor, the pressure in the compressor settles out to a level determined by a variety of factors. A blowdown valve is used to depressurize the compressor upon shutdown. Automatic control of the blowdown valve is recommended for high-risk locations and for compressors that are fitted with liquid film seals. When liquid film seals are employed, the compressor must be depressurized before the overhead seal tanks have been drained.
Discharge check valve
Placement of a check valve at the discharge of each section of compression can minimize or eliminate backflow through the compressor. Should backflow occur, it is possible for the compressor to experience potentially damaging reverse rotation. The presence of discharge check valves also provides the benefit of isolating each of the antisurge recycle loops (see the subsection on Antisurge Valves above).
The compressor develops its maximum pressure ratio when operating at both its maximum continuous speed and the surge control capacity. If the suction pressure increases for any reason, the discharge pressure correspondingly increases to the value given by the performance map for the speed and capacity in question. A pressure relief valve is installed to protect against overpressurization of downstream equipment by the compressor.
Before startup, it is necessary to purge air from the compressor and piping system. A purge valve is installed as a bypass to the suction shutdown valve for this purpose. Purging must be done with a low flow rate to prevent the purge gas from initiating compressor rotation. For this reason, the purge valve is small.
A discharge cooler (after cooler) is required if the temperature of the gas at the compressor discharge exceeds that required for the next step in the process.
Erosion of compressor components can be caused by ingestion of excessive liquid. To prevent erosion damage, suction scrubbers are installed to remove liquids that condense in the gas suction line because of cooling or that result from an upstream-process upset resulting in liquid carryover to the gas suction line.
A manual vent valve is installed between the compressor discharge and the discharge check valve to allow the compressor to be isolated from the vent header for maintenance. Once the compressor is shut down and blown down to the vent header, the blowdown valve can be closed and the vent valve opened. If the blowdown valve were kept open, there is a possibility that gas in the vent header would flow into the compressor system, endangering the maintenance operation.
Safety and monitoring devices
Centrifugal compressors are equipped with instrumentation to monitor mechanical health. Vibration monitoring is accomplished by eddy current probes installed at each of the compressor bearings. Vibration amplitude is measured at each radial bearing, and the axial position of the rotor is measured at the thrust disc or shaft end. The trend of radial vibration amplitude provides insight into the condition of the compressor regarding rotor balance and alignment. When a problem arises, the vibration frequency spectrum can also be analyzed to provide useful diagnostic information. The axial position probe monitors the state of thrust bearing wear. Each of the bearings is also fitted with temperature-sensing devices. By trending the thrust bearing pad temperatures, it is possible to discern the condition of the internal seals because changes in seal condition affect thrust loads and, therefore, bearing temperature. Alarm and shutdown settings for high bearing vibration and temperature are established in the compressor control system.
External to the compressor are numerous other alarm and shutdown safeguards. As a minimum, low lube oil pressure, low seal gas pressure differential, overspeed, high discharge gas temperature, high and low suction and discharge pressures, and high liquid level in the suction scrubber are monitored and will initiate a shutdown when necessary.
When properly designed, operated, and protected, centrifugal compressors are capable of long sustained runs with very little maintenance. The components most prone to wear are the bearing pads and internal labyrinth seals. Fouling of the internal surfaces can occur in some services causing a degradation of performance. The vibration and bearing temperature monitoring instrumentation, described in the section on safety and monitoring devices above, provides valuable information to the operator about the probable condition of the compressor bearings. Excessive wear of the internal labyrinth seals can occur when the compressor experiences high vibration excursions from process upsets or operation in a surge condition. Worn internal seals cause a degradation of compressor performance similar to that caused by fouling.
Unless there is an identified problem with the compressor, maintenance is generally carried out during planned turnarounds. As a minimum, the easily accessible compressor bearings and shaft endseals are inspected and replaced with spares, if necessary. Complete disassembly is required to inspect the compressor internals. The auxiliary lube and seal systems require maintenance of miscellaneous items such as seal gas filters, lube oil pump seals, oil filters, etc.
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