You must log in to edit PetroWiki. Help with editing

Content of PetroWiki is intended for personal use only and to supplement, not replace, engineering judgment. SPE disclaims any and all liability for your use of such content. More information


Reciprocating compressor

PetroWiki
Revision as of 12:20, 17 September 2013 by Glenda Smith (Glendasmith) (talk | contribs) (→‎Nomenclature)
(diff) ← Older revision | Latest revision (diff) | Newer revision → (diff)
Jump to navigation Jump to search

Reciprocating compressors are positive displacement machines in which the compressing and displacing element is a piston having a reciprocating motion within a cylinder. The discussion on the this page on reciprocating compressors includes a description of process configuration for multistage units, as well as an explanation of the concepts of:

  • Speed control
  • Inlet throttling
  • Recycling
  • Pressure relief
  • Blowdown
  • Distance piece venting and draining

Types of reciprocating compressors

There are two types of reciprocating compressors:

  • High speed (separable)
  • Low speed (integral)

The high-speed category also is referred to as “separable,” and the low-speed category also is known as “integral.”

The American Petroleum Institute (API) has produced two industry standards, API Standard 11P and API Standard 618, which are frequently employed to govern the design and manufacture of reciprocating compressors.

Separable compressors

The term “separable” is used because this category of reciprocating compressors is separate from its driver. Either an engine or an electric motor usually drives a separable compressor. Often a gearbox is required in the compression train. Operating speed is typically between 900 and 1,800 rpm.

Separable units are skid mounted and self-contained. They are easy to install, offer a relatively small initial cost, are easily moved to different sites, and are available in sizes appropriate for field gathering—both onshore and offshore. However, separable compressors have higher maintenance costs than integral compressors.

Fig. 1 is a cross section of a typical separable compressor. Fig. 2 shows a separable engine-driven compressor package.

Integral compressors

The term “integral” is used because the power cylinders that drive the compressor are mounted integrally with the frame containing the compressor cylinders. Integral units run at speeds of between 200 and 600 rpm. They are commonly used in gas plants and pipeline service where fuel efficiency and long life are critical. Integral compressors may be equipped with two to ten compressor cylinders with power ranging from 140 to 12,000 hp.

Integral compressors offer high efficiency over a wide range of operating conditions and require less maintenance than the separable units. However, integral units usually must be field-erected and require heavy foundations and a high degree of vibration and pulsation suppression. They have the highest initial installation cost.

Fig. 3 is a cross section of a typical integral compressor. Fig. 4 shows an integral compressor package.

Major components

Reciprocating compressors are available in a variety of designs and arrangements. Major components in a typical reciprocating compressor are shown in Fig. 5.

Frame

The frame is a heavy, rugged housing containing all the rotating parts and on which the cylinder and crosshead guide is mounted. Compressor manufacturers rate frames for a maximum continuous horsepower and frame load (see the section on Rod Load below).

Separable compressors are usually arranged in a balanced-opposed configuration characterized by an adjacent pair of crank throws that are 180 degrees out of phase and separated by only a crank web. The cranks are arranged so that the motion of each piston is balanced by the motion of an opposing piston.

Integral compressors typically have compressor and engine-power cylinders mounted on the same frame and are driven by the same crankshaft. Cylinders in integral compressors are usually arranged on only one side of the frame (i.e., not balanced-opposed).

Cylinder

The cylinder is a pressure vessel that contains the gas in the compression cycle. Single-acting cylinders compress gas in only one direction of piston travel. They can be either head end or crank end. Double-acting cylinders compress gas in both directions of piston travel (see Fig. 6). Most reciprocating compressors use double-acting cylinders.

Choice of cylinder material is determined by operating pressure. Cast iron is normally used for pressures up to 1,000 psi. Nodular iron is used for pressures up to 1,500 psi. Cast steel is usually used for pressures between 1,500 and 2,500 psi. Forged steel is selected for cylinder operating pressures greater than 2,500 psi.

A cylinder’s maximum allowable working pressure (MAWP) should be rated at least 10% greater than the design discharge pressure (minimum 25 psi). The additional pressure rating allows a high-pressure safety sensor (PSH) to be set above the design discharge pressure, and for a relief valve (PSV) to be set at a pressure above the PSH.

Wear compatibility of the rubbing parts (piston rings and cylinder bore, piston rod and seal rings, etc.) is also a criterion for selecting materials. Cylinders experience wear at the point of contact with the piston rings. In horizontal arrangements, cylinder wear is greatest at the bottom because of piston weight. Thermoplastic rings and rider bands are used in most reciprocating compressors to reduce such wear.

Cylinders are frequently supplied with liners to reduce reconditioning costs. Liners are pressed or shrunk in place to ensure that they do not slip. Replacement of a cylinder liner is much less expensive than replacing an entire cylinder. In addition, performance can be adjusted to new requirements by changing the inside diameter of the liner. However, cylinder liners increase the clearance between the valve and the piston, diminish the effectiveness of jacket cooling, and decrease compressor capacity from a given diameter.

Distance piece

The distance piece provides separation between the compressor cylinder and the compressor frame. Fig. 7 illustrates API Standard 11P and API Standard 618 distance pieces. Distance pieces can be contained in either a single- or double-compartment arrangement. In the single-compartment design, the space between the cylinder packing and the diaphragm is lengthened so that no part of the rod enters both the crankcase and cylinder stuffing box. Oil migrates between the cylinder and the crankcase. If oil contamination is a concern, an oil slinger can be provided to prevent packing lube oil from entering the compressor frame. For toxic service, a two-compartment design may be used. No part of the rod enters both the crankcase and the compartment adjacent to the gas cylinder.

The packing case should be vented to the first stage suction or a vent gas system. Distance pieces contain a vent to evacuate additional leaking process gas from the packing. The diaphragm and packing are designed to keep gas from entering the crankcase. Effective venting is required to ensure that the process gas does not contaminate crankcase oil.

Each compressor should be equipped with a separate vent and drain system for distance pieces and packing. Distance piece and packing vents should be piped into an open vent system that terminates outside and above the compressor enclosure at least 25 ft horizontally from the engine exhaust. The distance piece drain should be piped into a separate sump that can be manually drained. The sump should be vented outside and above the compressor enclosure. Lube oil from the sump can be mixed with crude oil or, under certain circumstances, must be transported for disposal or recycling.

Crankshaft

The crankshaft rotates around the frame axis and drives the connecting rod, piston rod, and piston (see Fig. 8).

  • Connecting rod connects the crankshaft to the crosshead pin
  • Crosshead converts the rotating motion of the connecting rod to a linear, oscillating motion that drives the piston
  • Piston rod connects the crosshead to the piston.

Piston

The piston is located at the end of the piston rod and acts as the movable barrier in the compressor cylinder. Selection of material is based on strength, weight, and compatibility with the gas being compressed. The piston is usually made of a lightweight material such as aluminum or from cast iron or steel with a hollow center for weight reduction. Thermoplastic wear (or rider) bands often are fitted to pistons to increase ring life and reduce the risk of piston-to-cylinder contact. Cast iron usually provides a satisfactorily low friction characteristic, eliminating the need for separate wear bands.

Wear bands distribute the weight of the piston along the bottom of the cylinder or liner wall. Piston rings minimize the leakage of gas between the piston and the cylinder or liner bore. Piston rings are made of a softer material than the cylinder or liner wall and are replaced at regular maintenance intervals. As the piston passes the lubricator feed hole in the cylinder wall, the piston ring gathers oil and distributes it over the length of the stroke.

Bearings

Bearings located throughout the compressor frame assure proper radial and axial positioning of compressor components. Main bearings are fitted in the frame to properly position the crankshaft. Crank pin bearings are located between the crankshaft and each connecting rod. Wrist pin bearings are located between each connecting rod and crosshead pin. Crosshead bearings are located at the top and bottom of each crosshead.

Most of the bearings in reciprocating compressors are hydrodynamic lubricated bearings. Pressurized oil is supplied to each bearing through oil supply grooves on the bearing surface. The grooves are sized to ensure adequate oil flow to prevent overheating.

Piston rod packing provides the dynamic seal between the cylinder and the piston rod. The packing consists of a series of non-metallic rings mounted in a case and bolted to the cylinder. The packing rings work in pairs and are designed for automatic wear compensation. Because each pair of rings accommodates a limited amount of pressure differential, multiple pairs are required depending on the pressure required by the application. To safely vent gas leakage through the packing, the vent port is usually located between the two outer ring assemblies (see the section on Distance Piece above).

Auxiliary connections to the packing may be required for:

  • Cooling water
  • Lubricating oil
  • Nitrogen purging
  • Venting
  • Temperature measurement

Lubrication must be finely filtered to avoid damage that would result from small particulate matter entering the case. The lubricating oil is normally injected into the second ring assembly, with pressure moving the oil along the shaft.

Compressor valves

The essential function of compressor valves is to permit gas flow in the desired direction and to block all flow in the opposite (undesired) direction. Each operating end of a compressor cylinder must have two sets of valves. The set of inlet (suction) valves admits gas into the cylinder. The set of discharge valves is used to evacuate compressed gas from the cylinder. The compressor manufacturer normally specifies valve type and size.

Plate valves constructed from rings connected by webs into a single plate are a common valve type. Depending on the sealing plate material, plate valves are capable of handling pressures as high as 15,000 psi, differential pressures to 10,000 psi, speeds to 2,000 rpm, and temperatures to 500°F. Plate valves do not perform well in the presence of liquids.

Concentric ring valves are capable of handling pressures to 15,000 psi, differential pressures to 10,000 psi, speeds to 2,000 rpm, and temperatures to 500°F. Advantages of concentric ring valves include:

  • Moderate parts cost
  • Low repair cost
  • The ability to handle liquids better than plate valves

Poppet-style valves generally provide performance that is superior to both plate and concentric ring valves. The poppet style uses separate, round poppets to seat against holes in the valve seat. This type of valve offers high lift and low pressure drop, resulting in higher fuel efficiency. Poppet valves are widely used in pipeline, gas conditioning, and processing facilities. Metallic poppets work well at:

  • Pressures to 3,000 psi
  • Differential pressures to 1,400 psi
  • Speeds to 450 rpm
  • Temperatures to 500°F

Thermoplastic poppets can be applied to applications with:

  • Pressures to 3,000 psi
  • Differential pressures to 1,500 psi
  • Speeds to 720 rpm
  • Temperatures to 400°F

Most compressors have valves mounted in the cylinders. A relatively new design concept places the valves in the piston. The valve-in-piston design (Fig. 9) operates with low valve velocities and provides longer life cycles and reduced maintenance time.

Compressor performance

Compressor capacity and horsepower are affected by piston displacement and cylinder clearance. The flow capacity of a given cylinder is a function of piston displacement and volumetric efficiency. Volumetric efficiency is a function of cylinder clearance, compression ratio and the properties of the gas being compressed. Compressor capacity may be calculated with any of the following three equations.

Vol3 page 287 eq 001.PNG................(1)

Vol3 page 287 eq 002.PNG................(2)

and

Vol3 page 288 eq 001.PNG................(3)

where

qa = inlet capacity of the cylinder at actual inlet conditions, Acf/min,
Ev = volumetric efficiency,
PD = piston displacement, Acf/min,
qg = inlet capacity of the cylinder, scf/min,
and
Qg = inlet capacity of the cylinder, MMscf/D.

Piston displacement

Piston displacement is defined as the actual volume of the cylinder swept by the piston per unit of time. Displacement is commonly expressed in actual cubic feet per minute (Acf/min). Calculation of the piston displacement is a straightforward procedure that depends on the type of compressor configuration. Single-acting cylinders can have either head-end or crank-end displacement. Eqs. 4 and 5 are used to calculate displacement of single-acting cylinders. Eq. 4 is for head-end displacement and Eq. 5 is for crank-end displacement.

Vol3 page 288 eq 002.PNG................(4)

Vol3 page 288 eq 003.PNG................(5)

where

PD = piston displacement, Acf/min,
S = stroke, in.,
N = compressor speed, rpm,
dc = cylinder diameter, in.,
dr = rod diameter, in.

Double-acting cylinder displacement is calculated with Eq. 6.

Vol3 page 288 eq 004.PNG................(6)

where

PD = piston displacement, Acf/min,
S = stroke, in.,
N = compressor speed, rpm,
dc = cylinder diameter, in.,
and
dr = rod diameter, in.

The methods used to change piston displacement include changing compressor speed, removing or deactivating suction valves in a double-acting cylinder, and changing the diameter of the cylinder liner and piston.

Unloading one end can significantly reduce the capacity of a double-acting cylinder. The best method to unload a cylinder is to deactivate or remove the suction valves from one end to prevent that end from compressing gas. Depending on the frequency of unloading and the molecular weight of the gas, a port or plug unloader is the next best method of unloading a cylinder. A doughnut replaces one suction valve of three or more valves per corner, and only one unloading device is required per cylinder end. With concentric ring-type valves, it is possible to place a plug unloader in the center of a suction valve for unloading. Depending on the molecular weight of the gas, both port and plug unloaders reduce BHP/MMscf/D and significantly improve the reliability of the unloading system.

If the suction valve is held open with finger depressors during the compression stroke, gas will flow through the open valve back into the suction gas passage, and no gas will be discharged from the end of the cylinder containing the unloaded suction valve. Deactivation of the valves may be performed manually while the compressor is shut down or by using a valve unloader or lifter while the compressor is in operation. Control of the valve unloader can be manual or automatic by a diaphragm that unloads the compressor using a suction pressure sensor. Diaphragm actuators are more reliable than manual lifters or unloaders.

Unloading both ends of the same cylinder may cause the cylinder to overheat; thus, it is best to unload only one end of a double-acting compressor cylinder. In most cases, it is preferable to remove the suction valve when unloading the head end of a cylinder to assure load reversal in the rods. (See the section on Rod Load below)

Clearance volume

Clearance volume is the space remaining in the compressor cylinder at the end of the stroke. Clearance is made up of spaces in valve recesses and the space between the piston and cylinder end. At the completion of each compression stroke, the compressed gas trapped in the clearance space expands against the piston and adds to the force of the return stroke. Fig. 10 is a pressure vs. volume (P-V) diagram illustrating the effect of clearance.

Expansion of the gas trapped in the clearance space occurs before the suction valve opens to admit new gas into the cylinder. As a result, part of the piston displacement occurs before the suction valve opens. The compression process in reciprocating compressors is nearly isentropic, so the energy required to compress the gas in the clearance space is recovered when the gas expands at the end of the compression stroke. For this reason, changes in the clearance space do not affect the compressor power.

Clearance volume is expressed as a percentage of piston-swept volume using one of the following configuration-dependent equations:

  • Single acting cylinder (head-end clearance) [Eq. 7]
  • Single-acting cylinder (crank-end clearance) [Eq. 8]
  • Double-acting cylinder (head-end and crank-end clearance) [Eq. 9]


Vol3 page 289 eq 001.PNG................(7)


Vol3 page 290 eq 001.PNG................(8)


Vol3 page 291 eq 001.PNG................(9)

where

%C = cylinder clearance, %,
CHE = head-end clearance, in.3,
CCE = crank-end clearance, in.3,
dc = cylinder inside diameter, in.,
dr = rod diameter, in.,
S = stroke length, in.

Application

Clearance can be added to a cylinder as:

  • Fixed-volume clearance pockets
  • Variable clearance pockets
  • Split-valve yokes
Fixed-volume clearance pockets

A fixed-volume clearance pocket is normally a volume bottle permanently attached to the cylinder. Fixed volume also can be added by a side-passage clearance plug consisting of a flange with a variable length plug inserted into a passage built into the side of the cylinder. A fixed-volume clearance pocket may be continuously open or may be controlled to be either open or closed. Control can be by manual hand wheel or automatic actuator. An actuator control allows the clearance pocket to be opened or closed from outside the cylinder while the compressor is in operation.

Variable clearance pockets

Variable clearance pockets allow a variable amount of clearance to be added to the cylinder and can be attached to either the head end or crank end of the cylinder. Most often, variable clearance pockets are attached to the head end, as shown in Fig. 11.

Split-valve yokes

Excessive clearance in a compressor cylinder can cause slamming of the discharge valves. If too much clearance is present, no gas will be discharged. Rapid overheating can occur because no cool suction gas enters the cylinder.

Volumetric efficiency

Volumetric efficiency is the ratio of the actual volume of gas (Acf/min) drawn into the cylinder to the piston displacement (cf/min). This ratio is less than unity because of three fundamental effects. First, the gas is heated during admission to the cylinder. Second, there is leakage past valves and piston rings. And third, there is re-expansion of the gas trapped in the clearance volume from the previous stroke. Of these three, re-expansion has, by far, the greatest effect on volumetric efficiency.

Compressor manufacturers have not reached consensus on an appropriate calculation method because measurement of these effects is extremely difficult. Recognizing this, the next approximate equation may be used to estimate volumetric efficiency.

Vol3 page 291 eq 002.PNG................(10)

where

Ev = volumetric efficiency,
R = compression ratio,
C = cylinder clearance, % of piston-swept volume,
Zs = inlet compressibility factor,
Zd = discharge compressibility factor,
dr = rod diameter, in.,
k = ratio of specific heats, Cp/Cv,
L = slippage of gas past piston rings, % (1% for high-speed separable, 5% for nonlubricated compressors and 4% for propane service),
and
96 = allowance for losses because of pressure drop in valves.

Rod load

Rod loads consist of gas loads caused by pressure and inertia loads that result from acceleration and deceleration of the piston, piston rod, crosshead, and approximately one-third of the connecting rod weight. Manufacturers specify a maximum rod load to protect the compressor because overloading the rods can severely damage the compressor. Loads must be evaluated for normal operating conditions and also during upset conditions. Rod loading must be reviewed at minimum suction pressure and relief valve pressure to assure an adequate safety margin.

Rod load reversals must be of sufficient magnitude to provide lubrication to the crosshead pin bushing. The bushings are lubricated by the pumping action of the opening and closing of bearing clearance that occurs when the rod load reverses from tension to compression. Operation without rod reversals also can severely damage the compressor.

Rod loads for the various compressor configurations are calculated with the following equations:

  • Single-acting cylinder (head end)
  • Single-acting cylinder (crank end)
  • Double-acting cylinder

Single-acting cylinder (head end)

Vol3 page 292 eq 001.PNG................(11)

Vol3 page 292 eq 002.PNG................(12)

RLc = rod load in compression, lbf,
RLt = rod load in tension, lbf,
ap = cross-section area of piston, in.2,
ar = cross-section area of rod, in.2,
Pd = discharge pressure, psia,
Ps = suction pressure, psia,
and
Pu = pressure in unloaded end, psia.

Single-acting cylinder (crank end)

Vol3 page 293 eq 001.PNG................(13)

Vol3 page 293 eq 002.PNG................(14)

RLc = rod load in compression, lbf,
RLt = rod load in tension, lbf,
ap = cross-section area of piston, in.2,
ar = cross-section area of rod, in.2,
Pd = discharge pressure, psia,
Ps = suction pressure, psia,
and
Pu = pressure in unloaded end, psia.

Double-acting cylinder

Vol3 page 293 eq 003.PNG................(15)

Vol3 page 293 eq 004.PNG................(16)

RLc = rod load in compression, lbf,
RLt = rod load in tension, lbf,
ap = cross-section area of piston, in.2,
ar = cross-section area of rod, in.2,
Pd = discharge pressure, psia,
Ps = suction pressure, psia,
and
Pu = pressure in unloaded end, psia.

Other performance factors

Additional performance considerations include:

  • Suction pressure. At constant discharge pressure with compression ratios greater than 2.0, the compression ratio decreases as the suction pressure increases. A decrease in compression ratio reduces the power requirement per unit of flow. The capacity of the cylinder, however, increases with suction pressure at a faster rate, resulting in an overall increase in power. To avoid overloading the driver, additional clearance must be added to reduce cylinder capacity.
  • Suction temperature. Cylinder capacity is inversely proportional to absolute suction temperature. As temperature decreases, more standard cubic feet fill the cylinder. Thus, a 10°F reduction in suction temperature increases compressor mass flow by almost 2%. Precooling the gas can be an effective way to increase cylinder capacity.
  • Discharge pressure. Changes in discharge pressure have little effect on cylinder capacity. Volumetric efficiency varies slightly with compression ratio, and the required power is directly proportional to the change in compression ratio.
  • Ratio of specific heats (k). An increase in k value produces an increase in the volumetric efficiency as defined by Eq. 10. Thus, a given compressor cylinder has a higher actual capacity when compressing natural gas (k = 1.25), compared with its capacity when compressing propane (k = 1.15). The higher capacity, when compressing natural gas compared to propane, results in greater power consumption as well.
  • Speed. Cylinder capacity is directly proportional to compressor speed. It is common practice to adjust compressor speed (within reasonable limits) to maintain desired suction pressure. Reduction of driver speed lowers fuel consumption and operating costs.

Performance maps

Performance maps can be developed for a specific compressor with base conditions held constant. Fig. 12 illustrates that as suction pressure increases, both inlet flow rate and power increase for constant discharge pressure and temperature. At very low ratios, the power may actually decrease with increasing suction pressure.

Process installation

The compressor is an integral part of a complete compression system. Fig. 13 is a typical process flow diagram for a reciprocating compressor installation.

Recycle valve

Compressor suction pressure decreases as the flow rate decreases until the gas expands to satisfy the flow required by the cylinder. The increase in compression ratio caused by reduction in suction pressure results in an increase in discharge temperature. Thus, the recycle valve in the system must be set to prevent low suction pressure from creating excessive discharge temperature. In addition, rod load limits may dictate the minimum acceptable suction pressure for a compressor installation. Where possible, recycle valve should be downstream of gas coolers.

Blowdown valve

The blowdown valve relieves trapped pressure when the compressor is shut down for maintenance. Valve control is typically automatic but is sometimes manual at some small, onshore compressor installations.

Suction scrubber

Ingestion of liquids into the compressor through the inlet gas stream can cause damage to the compressor internals. For this reason, an adequately sized suction scrubber with provisions for draining is required. The scrubber may be part of the pulsation control when properly planned (see section on Pulsation below). If the inlet stream is near saturation, horizontally-oriented cylinders and bottom-connected discharge nozzles are recommended.

Relief valves

Pressure relief valves set at a margin of 10% above the highest stage discharge pressure, or a minimum of 15 to 25 psi, provide static pressure protection for piping and coolers. Relief valve setting should never exceed the cylinder maximum allowable working pressure (see the section on cylinders above). Caution should be taken to ensure that all suction side gas piping, cylinders, and relief valves are rated for settle-out pressures in closed-loop refrigeration or low gas temperature services.

Pulsation

The flow of gas through a reciprocating compressor inherently produces pulsation because the suction and discharge valves are not open for the entire compression stroke. Pulsation damping is needed to create a more uniform flow through the compressor to assure uniform loading and to reduce piping vibration levels.

Pulsation control devices

If long, straight runs of piping of the same diameter as the compressor cylinder line connection can be provided, and the stage power is less than 150 hp, separate volume bottles or pulsation vessels may not be required. For most applications, volume bottles or pulsation vessels with internal baffles and/or choke tubes should be located as close to the cylinder as possible for optimum valve reliability. The addition of orifices at key locations in the piping can also reduce piping pulsations. Several different bottle-sizing formulae are available. Typical bottle sizes are five to ten times the cylinder swept volume.

Pulsation design

Digital piping pulsation analysis is a relatively low-cost method to ensure that a piping system is designed to meet acceptable pulsation levels (typically 2 to 7% peak to peak). The piping system layout must identify locations and volumes of knockout drums, bottles, coolers, and relief valves. The analysis should include the first major vessel or volume upstream and downstream of the compressor. Double-acting and single-acting (if applicable) operating conditions should be analyzed.

Vibration considerations

Imbalance of the rotating elements in the compressor cause mechanical vibration. Counterweights on the crankshaft and arranging cylinders in pairs on both sides of the crankshaft (in plain view) can minimize but not eliminate imbalance forces. Thus, there will always be mechanical vibrators that must be taken into account in foundation design.

Piping vibration

The compressor process gas piping must be properly designed and installed to avoid problems associated with excessive vibration. It is important that the natural frequency of all pipe spans is greater than the compressor pulsation frequency. The compressor pulsation frequency is calculated with Eq. 17.

Vol3 page 296 eq 001.PNG................(17)

where

fp = compressor pulsation frequency, cycles/sec,
N = compressor speed, rpm,
n = cylinder factor,
= 1 for single-acting cylinder
and
= 2 (for double-acting cylinder).

Piping should be securely tied using short pipe spans that are not uniform in length. Adequate pulsation damping helps prevent piping-related vibration problems.

Foundation design

For large integral compressors, or for compressors installed on complex structures or soft soils, it is best to perform a dynamic design using the imbalance forces provided by the manufacturer.

For high-speed compressors installed in areas with soils that can support a pickup truck, the following rules are useful.

  • Weight of concrete foundation should be at least three to five times the equipment weight.
  • Use soil bearing for a design that is less than 50% of that allowable for static conditions.
  • It is generally better to increase length and/or width rather than depth to meet weight requirements.
  • For rectangular block, at least 40% of height (but not less than 18 in.) should be embedded in undisturbed soil.
  • Concrete should be poured into a "neat" excavation without formed side faces.

Cylinder cooling

The heat of compression and friction between the piston rings and the cylinder add heat to the cylinder. Removing some of this heat is beneficial to the performance and reliability of the compressor in several ways. Cylinder cooling reduces losses in capacity and power caused by suction gas preheating. It also removes heat from the gas, thereby lowering the discharge temperature of the gas. Cylinder cooling also promotes better lubrication for longer life and reduced maintenance. When water is used as the cooling medium, uniform temperatures are maintained around the cylinder’s entire circumference, reducing chances for thermal distortion of the cylinder.

Care must be taken to avoid condensation that can result from excessive cooling. This can be assured by maintaining the cylinder jacket coolant temperature at least 10°F above the suction gas temperature.

Insufficient cooling can lead to reduced capacity and fouling of the cylinders. For this reason, it is recommended that the cylinder not be more than 30°F above the suction gas temperature.

Cooling systems

Types of cooling systems include:

  • Air cooled. Air-cooled systems are used for small throughputs and low heat loads. Cooling fins provide a sufficient surface area to cool the cylinder.
  • Static. Static systems are sometimes used on small compressors to assist air-cooled systems. Cooling fluid functions as a static heat sink and acts more as a heat stabilizer than a cooling system. Some heat is transferred from the system by conduction to the atmosphere.
  • Thermosiphon. The driving force for a thermosiphon derives from the change in density of the cooling fluid from the hot to cold sections of the system. API Standard 618 permits use of this system when discharge gas temperatures are below 210°F or when temperature rise across the cylinder is less than 150°F.
  • Pressurized. Pressurized cooling systems are the most common. In locations where cooling water is not available, a self-contained, closed cooling fluid system may be used. The system consists of a circulating pump, surge tank, and a fan-cooled radiator or air-to-liquid heat exchanger. The radiator may have multiple sections—one for cylinder coolant, one for cooling lube oil, and one (or more) for cooling discharge gas. The cooling fluid is either water or a mixture of water and ethylene glycol. The crankshaft usually drives the circulating pump.

Lubrication

Frame lubrication

The frame lubrication system delivers oil to the frame bearings, connecting rod bearings, and crosshead shoes. Some frame lubrication systems also supply oil to the packing and cylinders. For most reciprocating compressors, the lubrication system is integral with the frame.

Splash lubrication

Splash lubrication systems distribute lubricating oil by the splashing of the crank through the lubricant surface in the pump. Dippers may be attached to the crankshaft to enhance the effect. Splash systems are used on small, horizontal, single-stage compressors with power demands up to 100 hp.

The two main advantages of splash systems are:

  • Low initial cost
  • Minimal operator attendance

The main disadvantages are that splash systems are limited to:

  • Small frame sizes
  • The oil cannot be filtered

Pressurized lubrication

The most common type of frame lubrication is the pressurized system. Oil enters passages drilled into the crankshaft and flows through the main shaft and crank pin bearings. A pressurized lubrication system consists of the components discussed next.

Main oil pump

The main oil pump is driven by the crankshaft or may be separately driven. It is typically sized to deliver 110% of the maximum anticipated flow rate. When speed reduction is used for capacity control, care must be taken to ensure that this pump provides adequate lubrication at the minimum operating speed.

Auxiliary pump (optional)

An auxiliary pump is provided to back up the main pump. The auxiliary pump is usually driven by an electric motor and is designed to start automatically when oil supply pressure falls below a specified level.

Prelube pump (Optional)

A prelube pump supplies oil to the bearings before the compressor is started. This assures that the bearings are not dry at startup. Because this function is provided by the auxiliary pump, a prelube pump is required only when the system does not have an auxiliary pump.

Oil cooler

The oil cooler ensures that the temperature of the oil supply to the bearings does not exceed the maximum value required to protect the bearings from wear. A typical maximum oil supply temperature is 120°F. Jacket cooling water in a shell and tube heat exchanger is often used to cool the lubricating oil.

Oil filters

Oil filters protect the bearings by removing particulates from the lubricating oil. Some systems are equipped with dual, full-flow oil filters with transfer valves. Transfer valves allow switching from one filter to the other so that the filters can be cleaned without shutting down the compressor.

Overhead tank

The overhead tank provides oil to the bearings if a pump fails. The oil from the overhead tank is gravity-fed to the bearings. The tank must be sized to provide oil until the compressor has completely shut down. The tank is usually equipped with a level indicator.

Piping

The components of the lubrication system are connected by piping. Cleanliness and corrosion resistance are important considerations. Galvanized piping should be avoided because of possible corrosion. Carbon steel piping should be pickled or mechanically cleaned and coated with a rust inhibitor. Stainless steel piping should be used downstream of the filters. The piping system should be designed to avoid any pockets in which dirt or debris could accumulate. Socket welded piping should be avoided for this reason. Before initial startup, the lube oil system should be flushed with lube oil at approximately 170°F. A 200-mesh screen should be added to the system, and flushing should continue until the mesh is clean. Safety instrumentation should include a crankcase low oil level switch, a low oil pressure shutdown switch, and a high oil temperature switch.

For compressors with integral engine drivers, it is recommended that the compressor and driver be lubricated with separate systems so that combustion gases from the engine do not contaminate the lube oil. In this case, the packing and cylinder lubrication is provided by the compressor lubricating system. For installations in very cold environments, immersion or in-line heaters and special lubricating oils should be considered.

Cylinder and packing lubrication

The quantity of oil required to lubricate the packing and cylinders is small when compared with the bearing oil requirements. While the quantity is small, the oil pressure necessary to supply oil at the packing and cylinders is high. A small plunger pump (force-feed lubricator) is used at each stage of compression. Divider blocks are used to distribute the flow of oil between the cylinders and the packing. The oil can be supplied from either the frame lubricating system or from an overhead tank. Compatibility of the oil with the process gas must be checked to protect against contamination.


Nomenclature

qa = inlet capacity of the cylinder at actual inlet conditions, Acf/min,
Ev = volumetric efficiency,
PD = piston displacement, Acf/min,
qg = inlet capacity of the cylinder, scf/min,
Qg = inlet capacity of the cylinder, MMscf/D
PD = piston displacement, Acf/min,
S = stroke, in.,
N = compressor speed, rpm,
dc = cylinder diameter, in.,
dr = rod diameter, in.
%C = cylinder clearance, %,
CHE = head-end clearance, in.3,
CCE = crank-end clearance, in.3,
dc = cylinder inside diameter, in.,
dr = rod diameter, in.,
S = stroke length, in.
Ev = volumetric efficiency,
R = compression ratio,
C = cylinder clearance, % of piston-swept volume,
Zs = inlet compressibility factor,
Zd = discharge compressibility factor,
dr = rod diameter, in.,
k = ratio of specific heats, Cp/Cv,
L = slippage of gas past piston rings, % (1% for high-speed separable, 5% for nonlubricated compressors and 4% for propane service),
96 = allowance for losses because of pressure drop in valves
RLc = rod load in compression, lbf,
RLt = rod load in tension, lbf,
ap = cross-section area of piston, in.2,
ar = cross-section area of rod, in.2,
Pd = discharge pressure, psia,
Ps = suction pressure, psia,
Pu = pressure in unloaded end, psia
RLc = rod load in compression, lbf,
RLt = rod load in tension, lbf,
ap = cross-section area of piston, in.2,
ar = cross-section area of rod, in.2,
Pd = discharge pressure, psia,
Ps = suction pressure, psia,
Pu = pressure in unloaded end, psia
fp = compressor pulsation frequency, cycles/sec,
N = compressor speed, rpm,
n = cylinder factor,
= 1 for single-acting cylinder
and
= 2 (for double-acting cylinder)

References

Use this section for citation of items referenced in the text to show your sources. [The sources should be available to the reader, i.e., not an internal company document.]

Noteworthy papers in OnePetro

Use this section to list papers in OnePetro that a reader who wants to learn more should definitely read

External links

Use this section to provide links to relevant material on websites other than PetroWiki and OnePetro

See also

Compressors

Centrifugal compressor

Rotary positive displacement compressors

PEH:Compressors