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Theory and Application Of Reciprocating Compressors
Presented By George McKinney
Reciprocating/Positive Displacement Compressors Gas compression has been one of the anchor points of the industrial revolution, beginning with low pressure air supply for iron and steel refining, through higher pressure air supply for drilling and plant operating equipment, to high pressures as required for chemical synthesis, storage and pipeline deliveries of fuel gases. The positive displacement compressors in use today can trace their ancestry back to the original pumping machines invented by James Watts, or the bellows and blowers of blacksmiths. Piston type compressors have a solid position in this field: the technology is mature (more than a century of development), the fabrication process is straight forward, and the equipment is extremely scalable, ranging from miniature emergency tire inflation pumps to compressors of 10,000 horsepower or more. These latter are particularly used in the chemical process and gas transmission industries. There the requirements for high reliability, extreme range in throughput volume, and flexibility in operating pressures make an excellent fit for reciprocating piston compressors. This module describes the operating characteristics of various positive displacement compressors and develops the theory, basic calculations and rudiments of control for the piston type reciprocating compression process. While some references are to the gas compression and transmission industry, the same equipment construction and control methods are used in process compressors for industries such as petrochemicals and chemical synthesis.
1. Positive Displacement Compressor Types 1.1 Piston (Reciprocating) The reciprocating piston compressor is the most widely used equipment for gas service. The basic design consists of a piston in a cylinder with pressure actuated check valves to control suction and discharge flow through the cylinder. Standard practice is to have the piston driven by a rod passing through a packing case to seal against pressure leaks. With this double acting design, gas can be compressed on both sides of the piston. The basic design is more than a hundred years old, and is well developed. The throughput and loading can be adjusted by speed variation, addition of clearance to the cylinders, deactivating cylinders to reduce displacement or active control of valve closing, which effectively gives variable control of displacement. Efficiencies of this type of compressor can be more than 85 percent for conversion of horsepower input to pressure rise. 1.2 Vane A vane compressor consists of a cylindrical chamber with a rotating paddle wheel type drum mounted off center in the chamber. As the drum rotates, the sliding paddle wheel vanes section off volumes, which decrease in volume as they move toward discharge. A suction port is machined into the area where the chambers have the highest volume, and a discharge port is located where the chambers have the smallest volume. Gas enters at the higher volume and is compressed and discharged at the minimum volume. This type of compressor will tolerate more dirt than a reciprocating unit, and is often used for natural gas production services. The maximum differential is limited by the strength of the paddle wheel seals, so these units are not applicable for high pressures and differentials. 1.3 Blower (Rotary) In this compressor, two intermeshing elements rotate in an ellipsoidal chamber with intake and exhaust ports on opposite sides. As they rotate, gas is trapped in spaces formed between the chamber and moved to the opposite side of the chamber, where it is delivered to the discharge. This action is similar to the vane compressor, but is even more tolerant of liquids and dirt. For high pressure ratios, oil may be injected into the suction to improve the seal of the rotors and remove some of the heat of compression. 1.4 Screw(Rotary) The operation of a screw compressor is similar to the blower, except that the compression chambers are formed between two intermeshed elements similar to worm gears or screw threads. This compressor also requires oil injection for sealing and cooling. It is designed for high pressure ratios but is usually limited to discharge pressures below 250 Psig.
2. Reciprocating Piston Compressor Components A reciprocating piston compressor can come in two basic configurations. The simplest is a piston in a cylinder, directly driven from a crankshaft by a connecting rod attached to the piston by a wrist pin. This single acting (trunk type) piston can only compress gas on one face, and any leakage past the rings will vent into the crankcase. This can be hazardous with explosive, corrosive or poisonous gases, so this type of compressor is limited to applications where costs or simplicity are primary, such as shop air compressors. The illustration below shows a double acting compressor cylinder. In this case, the crankshaft drives a connecting rod which transmits force through a crosshead pin to a crosshead (similar to a trunk piston), moving in a slide. This converts the eccentric motion of the connecting rod to a pure linear force. A compressor rod connected to the crosshead transmits force to the compressor piston. In this case, the cylinder can be sealed on both ends, with the rod passing through a packing case to seal gas from leaking. This cylinder then can compress gas on both faces. By adding a vented space between the cylinder and crosshead, any leakage from the cylinder can be vented to a safe location, allowing handling of hazardous gases.
2A Engine Driven Double Acting Compressor Cylinder
2.1 Cylinder and Ends The compressor cylinder is a casting or forging designed to safely contain some maximum working pressure. It is machined to hold compressor valves and to direct gas flow to and from the cylinder cavity. In combination with the cylinder ends, it must contain the gas pressure, while having sufficiently large gas flow passages so there are minimal pressure drops due to gas flow. The cylinder and ends may also have water passages to stabilize temperature and dimensional changes. All these requirements involve compromises between size, strength, and flow passage size (efficiency). Compressor cylinders are designed for some operating range and service. If conditions change, they may not perform reliably or efficiently. As an example, a cylinder for gas transmission has large flow passages and valve areas for efficiency at high gas volumes and low pressure ratios, and will not function at high ratios. Similarly, a process cylinder may be a forging with small passages, giving higher strength but low efficiency. 2.2 Piston/Rings The compressor piston converts the energy/work supplied by the engine, applying it to the gas to raise its pressure. The piston must be strong enough to withstand the pressures and forces applied, but still be as light as possible, to minimize reciprocating weights and the resulting shaking forces. The compressor rings seal gas pressure to avoid leaking from one side of the piston to the other. The piston may also be fitted with a rider band, which is a low friction material to keep the metal piston from contacting the bore of the cylinder and causing scuffing and wear. Material for the rings and rider bands is selected to give long life and minimal wear with the typical pressures and gas composition of the compressor. While this is generally a low friction thermoplastic type material, rings may be made of bronze or other materials when temperatures are a problem. 2.3 Valves Compressor valves are simply fast acting check valves with a low pressure drop. They must be optimized to balance the opposing demands for long operating life and minimal pressure drop/flow losses. They may also have special features such as center ports to allow cylinder unloading. The compressor valve is possibly the most critical component when determining the requirements for a compressor service. The flow area is sensitive, as too small an area will give low efficiency, but too large an area can result in valve flutter and early failure. Similarly, valve components must be designed for the expected pressure and temperature conditions. Valves have been designed with many configurations, particularly in the sealing elements. These have progressed through steel, Bakelite, glass filled Teflon or Nylon, and high strength plastics. The most popular designs for sealing elements are ring shaped strips, mushroom shaped poppets, and straight channel strips.
2B Typical Compressor Valve Configurations - Cross Sections
Plate Type Valves/Single Deck
Poppet Type Valve/Double Deck
The design of compressor valves includes a number of variations to accommodate cylinder flow and unloading requirements. The simplest is a single deck valve, shown on the left above, where gas flows into passages in one face, across the sealing elements, and out through passages in the back face of the valve. A modification of this design is to have a large opening in the center of the valve. This allows adding a cylinder deactivator or clearance volume to the cylinder. This added feature comes at the expense of reduced flow area and efficiency. To compensate for this, two valves may be assembled together with a flow passage through the center. This double deck valve design has improved flow area and efficiency. This type of valve can only be used in a cylinder designed to accept its increased height. 2.4 Packing The compressor packing is a series of pressure containing rings located in the crank end of a double acting compressor cylinder. These seal against the piston rod and prevent leakage, so that the cylinder can compress gas on both sides of the piston. Again, as with compressor rings, the packing material is selected to provide best life and sealing with expected conditions. The packing is generally pressure lubricated, and may have coolant flow to remove friction heat. There are also various specialty types to reduce gas leakage around the rod. This may be important when compressing highly flammable or toxic gases. It is also becoming more important in reducing gas leakage and emission of “greenhouse gases”.
2.5 Clearance Unloaders In many applications, the volume of gas to be delivered may change based on either gas supply or process demands. Also, varying pressure conditions can change the load on the driver, requiring load control. This may be accomplished by speed variation, deactivating cylinders or cylinder ends, or by varying cylinder clearance. This last option is highly preferred, as speed control may have a limited range, and deactivating cylinders or ends can cause mechanical shaking or acoustic pulsations. Clearance unloaders allow varying throughput and load with minimal loss of efficiency. Unloaders are not actually a part of a compressor, but are included on many installations, to give load and throughput control. This may be done by volumes cast into the cylinder or heads, with a valve to close the passageway. Other options are valve cap pockets and head end variable pockets. Added clearance may have a simple handwheel to control its operation, or may have pneumatic actuators, which allow automatic operation. 2.6 Distance Piece Compartment(s) A distance piece section may be placed between the crosshead and cylinder to prevent leakage of gas from the compressor packing entering the compressor crankcase. At the crosshead end, an oil seal around the compressor rod prevents oil from migrating to the cylinder, and gas from entering the crankcase. This distance piece is normally vented to remove any gas which leaks from the packing. In cases of explosive or toxic gases there may be two distance pieces in series, to assure containment of the gases.
3. Definition of Terms 3.1
Single and Double Acting Compressor A Single Acting piston compresses gas on only one face, either by design or by deactivating valves on one side of a double acting cylinder Double Acting – Piston compresses gas alternately on both faces.
3.2
Connecting Rod A compressor element connecting the crankshaft to the compressor piston or crosshead. The connecting rod converts the rotation of the crankshaft into linear motion to drive the compressor piston.
3.3
Crosshead A crosshead is a sliding component at the outer end of the connecting rod, which converts the eccentric motion of the connecting rod to pure linear, eliminating side forces on the compressor piston.
3.4
Wrist Pin/Crosshead Pin
The wrist or crosshead pin connects the outer end of a connecting rod to either a single acting, trunk type piston (wrist pin) or to a crosshead (crosshead pin) 3.5
Compressor Rod/Piston Rod A cylindrical rod which connects the compressor piston to a crosshead, normally passing through a packing case to seal compression pressure into the cylinder
3.6
Compressor Piston A reciprocating component, normally fitted with piston rings which changes the volume of a cylinder, providing compression. It may be a simple trunk type piston directly connected to the connecting rod, or double acting, driven by a compressor rod.
3.7
Compressor Rings Compressor rings provide a seal between the compressor piston and cylinder wall, preventing gas leakage either into or out of the cylinder volume.
3.8
Rider Rings and Rider Bands Rider rings or bands are normally provided on a double acting piston to prevent contact of the piston with the cylinder wall. Rider rings/bands are normally made of carbon filled Teflon or other low friction material.
3.9
Compressor Packing Compressor packing is used in a double acting cylinder to seal around the compressor rod, preventing gas leakage from the cylinder. Packing is normally a series of segmented metallic rings, assembled and held in the end of the cylinder by the packing case.
3.10
Compressor Valves Compressor valves are high speed check valves, controlling flow of gas into the cylinder (suction valve) or out of the cylinder (discharge valve). They are designed for minimal pressure loss and maximum reliability
3.11
Cylinder Clearance (Mechanical) Clearance must be provided at the end of the piston stroke to avoid contact between the piston face and the compressor cylinder head. This clearance is expressed in linear measurement (inches or mm.).
3.12
Cylinder Clearance (Volume) Volumetric clearance is space left at the end of a piston stroke, both due to mechanical clearance and volumes above suction and discharge valves to allow for good gas flow. Clearance may also be added for control of throughput volume and/or load control (unloaders or clearance pockets).
This clearance is expressed as the ratio percentage of volume at the end of compression stroke to cylinder displaced volume. 3.13
Compression Ratio Compression ratio is the measure of increase in pressure across a compressor cylinder. It is determined by dividing the discharge pressure by suction pressure (both pressures must be absolute rather than gauge)
3.14
Pressure – Absolute and Gauge Gauge pressure is the value which would be measured by a gauge calibrated to indicate zero pressure when exposed to atmosphere. Absolute pressure is pressure which would be read from a gauge calibrated to read zero when exposed to complete vacuum. Normally absolute pressure is gauge pressure + 14.73 PSI.
4
Reciprocating Compressor Theory
Cycle Events In a reciprocating compressor, the process follows four main events – compression, discharge, re-expansion and intake. The first two are accomplished as the piston moves forward, reducing cylinder volume, while the second takes place as the piston moves back down the cylinder. For a more complete picture, assume starting the cycle with the compressor at the bottom of its stroke, with maximum cylinder volume. The cylinder is full of gas at suction pressure, and both suction and discharge valves are closed by gas pressure. As the piston moves forward, the cylinder volume decreases and pressure rises. When the cylinder pressure rises slightly above discharge pressure, the discharge valve opens and gas is pushed into the discharge piping for the rest of the stroke. At top center, the discharge valve closes. As there must be clearance between the piston face and cylinder head to prevent parts hitting each other, some volume of gas is trapped in the cylinder at discharge pressure. As the piston moves back down the cylinder, this gas re-expands until it reaches suction pressure. At this point, the suction valve opens and a fresh charge of gas flows into the cylinder for the remainder of the stroke.
4.1 Volumetric Efficiency As noted above, the cylinder does not bring gas in through the entire piston travel. The percentage of stroke the suction valve is open, compared to the entire stroke is called “volumetric efficiency”. If there were no clearance (volume) left when the piston completed its compression stroke, then cylinder pressure would immediately drop to suction pressure as the piston returned, giving 100 percent volumetric efficiency. Thus, the cylinder displacement would be equal to the volume delivered with each stroke. However, due to gas re-expansion, the suction valve opening is delayed. This delay becomes greater when the cylinder pressure ratio increases or the clearance volume increases. Thus, the cylinder delivers a reduced volume to the discharge condition. The pictures below illustrate this effect, with the picture on left showing effect of increasing clearance, and on right the effect of increasing pressure ratio. At high pressure ratios, or with large amounts of clearance, the valve opening may be delayed to the point that the valve does not open, and no gas flows through the cylinder. This condition is called zero volumetric efficiency, and can cause serious cylinder heating problems. In normal operation, friction of rings on the cylinder creates heat which is carried away with the gas being compressed. Since at zero volumetric efficiency, no gas is entering or leaving the cylinder, all friction heating effects are contained within the cylinder, causing an uncontrolled temperature rise. As the hot gas is contained within the cylinder, normal temperature detection in the discharge line will not be effective.
Effect of ratio and clearance on Volumetric Efficiency
4.2 Clearance Control As noted above, cylinder clearance will significantly affect throughput and horsepower of a compressor. Some amount of volumetric clearance is built into the cylinder to prevent the compressor cylinder from contacting the heads at the extremes of piston travel, and to provide a smooth gas flow path into and out of the cylinder. Beyond this, additional clearance can be introduced by providing clearance pockets or passages which open into the cylinder cavity. These have valves which can be opened or closed to add or remove the clearance from the compression process. Also, some cylinders may be equipped with a variable clearance pocket on the outboard cylinder head. These have a piston positioned by a screw and hand wheel, which will add a variable amount of clearance. 4.3 Work of Cycle The familiar definition of work is force times distance. In the pressure-volume cards shown above, piston movement or change in volume defines a distance. As the force against the piston changes as pressure increases and decreases, the area of the card defines the work involved in the cycle. A key point to note is that for a given pressure differential, changing the volumetric efficiency changes both the volume delivered and the work of the cycle. This is the basis for load control of compressors by changing the cylinder clearance. 4.4 Pressure Ratio Pressure ratio is the discharge pressure of the compressor divided by the suction pressure. These pressures must be in absolute (Psia) rather than gauge (Psig) pressure. As most operating gauges read in Psig, atmospheric pressure must be added. This is normally about 14 Psi. A reciprocating compressor may be able to operate at high pressure ratios, but is usually limited by other conditions, particularly temperature. A compressor’s discharge temperature increases with pressure ratio. For example, at a pressure ratio of four and a suction temperature of 60 degrees, discharge temperature would be about 310 degrees. This is a safe practical limit for most compressor components. Consequently, pressure ratios across any single compressor cylinder rarely are allowed to exceed four to one. 4.5 Temperature Rise – Ratio Effect When a gas is compressed, its temperature rises in proportion to the pressure ratio. For low pressure ratios, the discharge temperature may be only twenty to
fifty degrees higher than suction temperature. When the pressure ratio is high, such as on storage or production service, the discharge temperature may be more than a hundred degrees higher than the suction. This is true for all types of compressors. This temperature rise may limit the amount of pressure rise allowable across a compressor, or require special components to withstand the temperature. This temperature must be reduced before gas is put into underground pipelines, to prevent melting their protective coatings. In most cases, the discharge temperature from a compressor station must be kept below 1250F, requiring gas coolers at higher pressure ratios. This is particularly the case at storage and production stations, where high pressure ratios give extreme discharge temperatures.
5 Compressor Operating Characteristics When installing or operating a compressor, it will help to understand the reasons for selecting a particular compressor type and its optional equipment. The following describes some of the characteristics of reciprocating compressors and the need for various features. Compressor Limitations 5.1 Working Pressure Compressors are designed for a maximum stress on the cylinder body and the attachment of heads on both ends. This maximum working pressure must not be exceeded. 5.2 Temperature Compressor discharge temperature is a function of pressure ratio; as the pressure ratio rises, discharge temperature rises also. Maximum discharge temperature will be limited by the materials in the compressor valves, rings and packing. Most commonly used materials have a limit of 250-275 degrees. High temperature materials, such as PEEK or steel valve plates will allow operating at discharge temperatures up to 350 degrees. 5.3 Compressor Rod Loading In a double acting compressor, the piston rod receives the force of gas pressures acting against the piston. The head end produces a compressive force, equal to the pressure in the cylinder multiplied by the area of the piston. The crank produces a tension force, again equal to pressure times area of the crank end of the piston. These forces vary as the piston moves from suction to discharge events. As piston area is constant for the two faces, the rod loading can be expressed as: (5.3a) (5.3b)
Compression = (Ah*P2 –Ac*P1) and Tension = (Ah*P1 – Ac* P2) Where Ah = Area of outboard (head) end of piston, sq. in. Ac = Area of inboard (crank) end of piston, sq. in. P1 = Suction pressure, Psig P2 = Discharge pressure, Psig
From the above equation, it can be seen that for some maximum value of compression or tension loading, there will be a maximum differential of suction to discharge pressure.
In normal operation, as the piston goes through discharge on the outboard side, it will have its maximum compressive force on the rod, and when discharging on the inboard event, will give the maximum tension force. The sum of the head and crank forces must be kept within a limit established by the manufacturer. This represents the limit of strength of some component, either attachment of the rod to the piston or crosshead, or strength of the oil film at the crosshead pin bushing. In addition, the rod loading should reverse from compression to tension for some specified period, to allow the oil film to rebuild at the crosshead pin and bushing, preventing loss of lubrication and early failure. 5.4 Volumetric Efficiency Volumetric efficiency is normally expressed as a percentage of the compressor stroke where a valve is open. In almost all cases, it is stated as a percentage of suction volumetric efficiency. The accompanying graph shows actual test points of suction volumetric efficiency for a single stage storage unit. The points form a definite line for each of the clearance conditions, with all the lines intersecting at a pressure ratio of 1.0 and a volumetric efficiency of 100 percent. This shows the effects of ratio and clearance on volumetric efficiency. By extending the plotted line and ratio scale, it can be seen that for any clearance condition the line would go to zero at a high ratio. In a compressor with good design, clearance is balanced as much as possible over all cylinder ends. When unloading the engine, the clearance can then be spread over all cylinders, which keeps the volumetric efficiency as high as possible. In the case of high speed compressors, this may be impossible due to cylinder design.
5A Typical Observed Volumetric Efficiency
5.5 Flow Reduction/Volume Control As noted earlier, compressor flows can be changed by either speed variation or changing the compressor cylinder clearance. Speed control will have a direct effect at any condition. Therefore, a ten percent speed change would have a ten percent effect in flow. Clearance changes will have a varying effect depending on the pressure ratio of the compressor. At low ratios, clearance will have little effect. At high ratios, clearance will have much more effect on throughput. Effect on Engine/Driver Flow changes have a direct effect on the driver. If the compressor flow is reduced by ten percent and the pressures do not change, the engine load will be reduced by ten percent also. If we reduce the volume by reducing engine speed, engine horsepower is reduced, but engine torque will be relatively unchanged. As engines develop their best fuel efficiency at peak torque, reducing speed to drop throughput provides the best economy. However, this approach is limited by minimum equipment operating speeds. Adding clearance reduces volume (and horsepower) while keeping the speed constant. Therefore, adding clearance reduces both torque and horsepower. Speed variation is more often used for flow control, while clearance is used primarily for engine load or torque control. 5.6 Compression Efficiency - Ratio Effect In the previous section, the work involved in increasing gas pressure was evaluated from suction to discharge conditions, with no pressure drops assumed. Actually, there are pressure drops involved in bringing gas into and out of the cylinder through gas passages and compressor valves. These losses create additional work in the cycle but do nothing for increasing the effective pressure differential. The efficiency of a compressor is the percentage ratio of useful work done in raising the gas pressure to the total work supplied. At low pressure ratios, the effective work in raising gas pressure is low, while throughput of the cylinder is usually high, giving high velocities in gas passages and through compressor valves. Thus, much of the work supplied goes into moving the gas through cylinder valves and passages rather than actually increasing the pressure. At high ratios, more work is required to increase gas pressure and proportionally less is wasted in flow losses, giving high efficiency. 5.7 Range With the above points, it can be seen that compressor efficiency is low at low ratios, where flows are high and most of the work goes into moving the gas
through the cylinder. In fact, at a ratio of one (suction equal to discharge), efficiency goes to zero, as no useful pressure is built. As the ratio increases, the efficiency rises, typically to a maximum around 85-90 percent on low speed units. This efficiency illustrates some of the design compromises of compressors. A process compressor will be inefficient at low pressure ratios, being designed for high ratios. A transmission compressor will be more efficient at lower ratios, but its design is not acceptable for higher ratios. 5.8 Speed Effect As noted above in flow effects, losses are higher with increased flow due to increased velocity and pressure drop through passages and valves. If the operating speed of the compressor is reduced, there is more time allowed for a given volume to flow through restrictive elements, so the flow losses will drop. This results in an increase of efficiency when the speed of a compressor is reduced. Typically, a compressor’s efficiency is evaluated at maximum speed. Then, if speed is reduced, the efficiency increase will result in the engine being slightly underloaded. Another option is to make a speed correction for efficiency.
5B Typical Observed Compression Efficiency
5.9 Low/High Speed Compressors Cylinder Design The basic design of compressor cylinders as outlined above is common for all units. The chief difference between low and high speed cylinders is due to the length of stroke. While low speed units have stroke lengths from 14 to 20 inches, a high speed compressor’s stroke can range from 3 to 7 inches. Because of the
shorter stroke in high speed units, a large part of the valve(s) is covered by the piston at the ends of the stroke, impeding gas flow and reducing the effective valve area. Valve Efficiency Because of the cylinder design with valves being covered by the piston, a high speed cylinder is normally five to ten percent lower efficiency than low speed cylinders. This is compounded by the difficulty in building high valve element lifts and large flow areas into valves running at high speeds. Unloader Capability In both low and high speed units, the simplest option for load control is to add clearance on the outer (head) ends. For more flexibility, clearance may be added on both ends. In some cases, this is done by means of clearance passages, which are holes passing through the cylinder wall allowing clearance to be added externally. Another option is by adding valve cap clearance pockets. In this case, a valve will have a hole through its center to allow free flow of gas from the cylinder into the clearance pocket. As this reduces the flow area and efficiency of the valve, a double deck valve is often used to restore the flow area. For this, a deep valve cavity is needed to accept the increased valve height. Generally, high speed cylinders are not designed with these deep pockets, as it makes a larger outer diameter for the cylinder. This would be harder to fit on the small high speed compressor frame.
Rod Loading Capability In a compressor, the driving force of the engine is transmitted through the crank throw to a connecting rod and crosshead assembly to a compressor rod which has the piston attached. The driving force of the engine is being countered by the pressure of the gas being compressed, acting against the faces of the piston. This balancing of forces acts through a number of threaded connections and bearings with oil film lubrication. The compressor rod load represents a mechanical limit, beyond which some of these components can fail. Because of the longer stroke and heavier components of low speed units, it is easier to obtain high rod load capabilities. Higher speed units cannot tolerate high reciprocating weights, and so their compressor rods and bearing surfaces are proportionately smaller. This results in lower rod load ratings.
6
Pulsation Characteristics
Pressure pulsations are created by rapid variations in pressure. If these variations are repeated at some definite frequency, they can build energy, resulting in high levels of pulsation. These can cause loss of efficiency, piping movement, inaccuracy in measurement and eventual equipment failure. At the compressor level, this may appear as compressor valve failures or inaccuracy in prediction and control of throughput and engine load. 6.1 Generation The design of a reciprocating compressor results in a pulse on the suction and discharge side of piping each time the valves open. Because most compressors are double acting, there are two pulses generated for each revolution of the crank. With multiple compressor cylinders, there will be pulses created for each, with a definite phase timing related to the attachment of the compressor to the crank. This series of repeated pulses is fed through the suction and discharge piping system, and can cause shaking if some piping components are resonant at the frequency of the pulsations. 6.2 Filtering Pulsation filtering is generally done by either providing a large volume on the suction and discharge bottles or by creating a piping filter system which is tuned to the most critical pulsation frequency. The first method is simpler, and will attenuate all frequencies to some degree. However, it may result in unacceptably large volumes. The second method requires an engineering study of the compressor and associated piping. This is generally more expensive than the simple volume approach, but is capable of predicting and greatly reducing problem levels of pulsation. 6.3 Effect on Compressors If pulsations reach high levels during times when compressor valves are open, valve plates may flutter or their closing may be delayed, resulting in valve plate breakage. High pulsation levels may also cause early or delayed valve opening, causing unpredictable flows and horsepower levels. There may also be excessive unbalanced forces from end to end of pulsation bottles, resulting in high levels of vibration and possible cracking of piping.
7
Multi-Staging
7.1 Sharing Differential The limits of operation listed above show that a reciprocating compressor has a number of mechanical limits, most of which are related to pressure differential. Often differentials are required greater than can be accomplished with a single stage of compression. In this case, it is necessary to have multiple stages of compression. This is accomplished by having a cylinder or cylinders which take gas in at a low pressure, compress and discharge to an intermediate pressure, then repeat with additional cylinders to take the gas to the discharge pressure. In this process, pressure differential and temperature rise across each cylinder can be controlled to a reasonable level. The gas may be cooled between stages to minimize discharge temperatures. Normally this is done with two or more cylinders on the same compressor unit, with gas cooling between stages. 7.2 Efficiency Increase When gas is compressed, the temperature rise effectively creates higher volume at the discharge conditions. This requires more energy (work) for compression. In multiple stage compression with cooling, the temperature rise is minimized, which reduces the total work required to compress to the final discharge. 7.3 Operating Difficulties Multiple stage compression presents challenges for both design and operation. At the design stage, cylinders must be sized so that all stages are operating within their limits. In operation, the pressure balance between stages must be maintained by following a specified unloading sequence when pressures change, or when controlling engine load. Mechanical failures such as leaking compressor valves or rings can cause pressure unbalance, which may put excessive differentials or temperatures on other stages. The compressor piping and pulsation bottles will also be more complex, which will probably require an electric analog or digital evaluation to avoid pulsation or vibration problems.
8
Compressor Control Systems
8.1 Horsepower Requirements In general, compressor horsepower requirements increase as pressure ratio increases. The horsepower also increases as flow increases. But as the ratio increases, volume decreases. These characteristics act in opposite directions as suction is varied assuming a constant discharge pressure. The combination results in a general compressor characteristic of increasing suction causing increasing load. Over a wide suction range, load will increase, reach a peak and then decrease. Compressor design attempts to provide adequate piston displacement to load the engine at the minimum load points. The design must then incorporate adequate load control provisions to keep the engine in an acceptable condition at the maximum load point. 8.2 Clearance Volume Controls Load on a compressor unit can be controlled by adding or taking away clearance volume on the compressor cylinders. This reduces the cylinder’s volumetric efficiency, effectively reducing displacement in small amounts. This is normally the preferred method of controlling load. Problems with this approach are added cost and physical size and limitations of added clearance volumes. If the unloading is not evenly distributed among all cylinder ends, it is possible to have the ends with more unloading stop pumping. This may result in extreme temperature buildup in the affected ends. 8.3 Deactivation Engine load may also be controlled by deactivation of compressor cylinder ends. If a compressor has four double acting cylinders, each cylinder end is absorbing about one eighth of the total horsepower. Deactivating a cylinder end would reduce engine load by one eighth. This may create problems, in that the horsepower reduction may be more than desired. Also, when a cylinder is deactivated, gas continues to be pulled in, then pushed back into the suction. This wastes some horsepower, and heats the gas, which is then compressed by the other cylinder ends. This will result in higher discharge temperatures. Deactivation will also introduce odd harmonic pulsations into the discharge piping, which may cause piping shaking.
8.4 Active Control Systems Some compressor units have been equipped with load control systems to partially deactivate the compressor cylinders. In this, a device holds the suction valve open for a part of the compression stroke, allowing gas to flow back, as with a deactivated cylinder. At some point in the compression stroke, the valve is allowed to close and compression and discharge occur in a normal pattern. In this way, load and throughput can be controlled in very small increments. This system requires a control unit to time the valve closing and to send signals to a valve lifter device. The actuation is normally hydraulic, requiring a separate pump, control valves to supply oil to each compressor valve, and tubing. This type of system can provide great flexibility in throughput and loading. Its disadvantages are mechanical complexity and some loss of efficiency due to gas being pushed back through the valve while it is held open.
9
Compressor Torque Control/Throughput Control Systems
9.1 Throughput Control In some cases, the only consideration is to regulate throughput of a compressor unit or station. For this, control may be quite simple. As an example, with a given discharge pressure and constant speed of the compressor, volume will decrease as the suction pressure decreases. For this, a suction pressure regulator may be sufficient to control throughput. This will allow for variations of pressure, such as may be seen in a production field, while keeping the engine at a relatively constant load and throughput. The same principle can be extended to control of either engine speed or of a suction controller to maintain suction or discharge pressure. These control systems are simple closed loops, where an offset from some setpoint causes a feedback, which generates a control output to restore conditions. This is a simple analog control, with minimal logic and computation. It also usually assumes that there is sufficient horsepower installed to operate safely at any condition which may be allowed. Because of this, the compressor will often be either operating at less than its full capability, or with lowered efficiency due to regulation of pressures. 9.2 Torque Control – Unit Optimization Another option to control a compressor unit’s load or throughput is to use some of the control variations listed above, along with a computerized system to calculate the unit’s operation. This has the advantage of providing optimum operation of the engine/compressor, and maximum throughput capability for the installed
horsepower. Its disadvantage is complexity of equipment and need to have accurate prediction methods for calculating compressor throughput and load. There are two basic approaches to calculating and controlling engine load and throughput. One is to measure engine parameters and infer compressor operation, while the second is to measure pressures and calculate compressor performance, then assume engine output. 9.3 Engine Parameter Calculation This is an advanced application of the “closed loop” approach noted above. An engine in good condition will have a definite relationship between the amount of horsepower generated and its fuel requirement. Thus, a calculation of developed horsepower can be made from fuel measurement. This may be in terms of fuel flow (volume) or fuel pressure downstream of the governor. In the case of a four-cycle engine, horsepower is also related to intake manifold vacuum. These approaches have the advantage of controlling based on a readily available engine parameter. The disadvantage is that they assume a properly functioning engine. A misfiring power cylinder or improperly adjusted air/fuel ratio will give erroneous results. The benefit is that most errors will be in the direction of increasing fuel usage, giving an indication that the engine is developing more horsepower than actual. Thus, the system acts to protect the engine in most cases. 9.4 Compressor Calculation A compressor usually provides stable and easily monitored conditions for calculations of throughput and developed engine load. These calculations can be adjusted for the effect of the various methods of load control. The accuracy of the prediction and load control is easily established, as standard maintenance or performance analyzers will provide output information, which can be directly compared with the control program. The disadvantage of this type of control is the assumption that the compressor and engine are in good condition. If the engine is in poor condition, the compressor calculation will provide accurate loading, but the engine may not be in shape to maintain its rated output. Also, some compressor related problems may lead to overloading. Some conditions that can cause overloading are buildup of fluids in unloader pockets and accumulation of dirt or salt on valves and passages. This can result in restrictions which reduce the efficiency of the compressor.
10 Basics of Compressor Calculations All reciprocating compressor calculations are based on the compressor’s characteristics and operating conditions. For this, we need to know: Compressor physical description – bore, stroke, rod diameter and number of cylinders. This also includes number, location and size of any clearance type unloading provisions. Compressor operating conditions – suction and discharge pressures and any pressure drops from the sensing point to the compressor cylinder. Compressor running speed and status of operating cylinders. This includes any deactivated cylinders or ends and any added clearance volumes. Gas Calculation factors – In all calculations of gas conditions, pressures are normally measured in absolute values. This is gauge pressure with atmospheric pressure added. In the same way, temperatures are usually measured in degrees Rankine. A close approximation is to add 460 to the Fahrenheit reading to convert. Both of these corrections are made so that calculations refer to pressure and temperatures above the point of absolute zero temperature and pressure. At the root of all compressor calculations is the suction to discharge pressure ratio across the machine. This is based on the pressures inside the compressor, so any pressure drops from the compressor to gauge readings must be included. It is also calculated from absolute pressures, so the atmospheric pressure is added to gauge pressures. As an equation, it can be expressed: (10.1)
Rc = (P2 + DP2 + Atm)/(P1 + DP1 + Atm) Where P1 = Suction Gauge Pressure P2 = Discharge Gauge Pressure DP1 = Suction Drop, Gauge point to cylinder DP2 = Discharge Drop, Gauge point to cylinder Atm = Atmospheric pressure, normally 13.2 – 14.7 Psi
Knowing the pressure ratio across a compressor, the discharge temperature can be calculated with the following formula: (10.2)
T2=T1*(Rc(K-1)/K) Where T2 = Discharge temperature (Rankine) T1 = Suction temperature (Rankine) K = Gas ratio of specific heats (normally 1.2-1.3 for natural gas)
The compressor volumetric efficiency is the next level of calculation necessary for any prediction. As noted above, it is a measure of the effective displacement of the compressor, as opposed to the actual displacement due to the piston’s movement. When the piston reaches the end of its stroke at Top Dead Center, the gas remaining in the cylinder re-expands as the piston moves back down the bore. This delays the point where the cylinder pressure drops below suction, allowing a new charge of gas into the cylinder. This effect is a function of the volume of gas trapped at the end of the stroke and the pressure ratio. Volumetric efficiency decreases as the cylinder pressure ratio increases and as the clearance volume increases. This can be expressed as: (10.3)
Ev = 1- Cl* (Rc1/K-1) Where Ev = Volumetric Efficiency as a decimal percentage Cl = Compressor average clearance percentage (Clearance cubic inches/Displacement Cubic inches) Rc = Cylinder Pressure Ratio K = Gas ratio of specific heats
This formula will frequently have additional correction factors. In many cases, a slippage factor will be subtracted to adjust for cylinder leakage effects. In addition, the ratio factor may be multiplied by a ratio of supercompressibility factors to correct for non-ideal compression and re-expansion of the gas. Compressor delivered volume is based on displacement, volumetric efficiency, and gas conditions. The equation for capacity can be developed as: (10.4) Actual Displacement, CFM = PD*RPM * Ev Where PD = Cylinder displacement in Cubic Ft. RPM = Compressor running speed Ev = Volumetric Efficiency (percent) This would provide the actual volume of gas being moved. However, gas is measured and sold at standard conditions. This is defined as an absolute pressure of 14.73 Psi, and temperature of 60 degrees F. To represent the volume being moved in standard units, we multiply by the ratio of suction pressure (absolute) to standard pressure, and standard temperature to suction flowing temperature. In this calculation, the temperatures must be expressed in Rankine degrees (Fahrenheit + 460).
Also, as noted above, the gas volume must be corrected for supercompressibility. This is an experimentally determined adjustment for the non-perfect relation of pressure to volume as gas is compressed. The effect is that at higher pressures, more molecules of gas can fit into a volume than would be the case for an ideal gas. While gas volumes are measured at elevated pressures, the gas is bought and sold based on standard base conditions, typically 60 degrees F. temperature and 14.73 Psi (Absolute) pressure. The supercompressibility correction must be made to relate to these standard conditions. So, the equation becomes: (10.5)
Volume(Standard Cubic Ft./Minute) = PD*RPM*Ev*((Ps+14.73)/14.73)*((460+60)/(Ts+460)*1/Zs Where Ps = Suction pressure at the cylinder Ts = Suction temperature Zs = Supercompressibility at suction conditions
This needs only to be corrected for units. The standard measurement of volume is in millions of standard cubic feet per day. So, by multiplying by 1440 minutes per day and dividing by one million, we have the final equation: (10.6) Capacity(MMSCFD) = PD*RPM*Ev*(Ps/14.73)*(520/Ts)*1/Zs*1440/106 Horsepower Requirement- knowing the throughput volume of a compressor and its suction and discharge conditions, the horsepower required for compression can be calculated. This is represented by: Horsepower = Capacity (MMCFD)* Hp/MMCFD*1/Ec*1/Em. In this equation, capacity is the equation derived above without the correction factors for temperature and supercompressibility. Hp/MMCFD is the energy requirement to raise one million cubic feet of gas from the suction to discharge condition. This is actually adiabatic Hp/MMCFD, where the compression is assumed a perfect process, with no heat being transferred. This means that during the compression process, no heat is absorbed by the cylinder wall or compressor piston, and no heat is radiated into the gas. The formula for this is: (10.7) Bhp/MMCFD = 43.636* K/(K-1)*(Rc (K-1)/K –1)* (Zs+Zd)/2Zs 10.8 Compression Efficiency (Ec) is the percentage of supplied energy that actually goes into raising the pressure of the gas. The adiabatic horsepower is based on a theoretically ideal cycle, with no losses. In an actual cycle, there are energy losses due to flow losses across the compressor valves and cylinder
passages. There are also losses due to pressure pulsations. These reduce the compression efficiency, with the worst losses occurring at lower pressure ratios. The extreme case is where suction and discharge pressure are equal (compression ratio equals 1), when all input work is wasted in flow losses, with no increase in discharge pressure (no effective work). This results in compression efficiency values starting at zero at a ratio of 1.0 and increasing to peak values in the range of .78 -.92 at ratios of 2.0 or higher. 10.9 Mechanical Efficiency (Em) is a factor to correct for mechanical friction in the compressor. This covers bearing friction and the friction of compressor rings and packing. From extensive manufacturer’s testing, this is assumed to be .95 for most large integral compressors. On high speed and separable units, it is often assumed as .93 or less, due to the added friction losses of a separable unit’s crankshaft and other components. Combining the capacity and Bhp/MM equations results in a final horsepower equation of: 10.10 BHP = 43.636*.001*Pd*Ev*P1*(K/(K-1))*(RcK-1/K-1)*1/Ec*1/Em*(Zs+Zd)/2Zs From this, several things can be seen: 1. Horsepower and capacity are directly related – as capacity increases or decreases, horsepower will also change in the same proportion. 2. Change in volumetric efficiency is one key point to control both capacity and horsepower. 3. Changing piston displacement will also have a direct effect on capacity and horsepower. 4. Compression efficiency has no direct effect on capacity, but can have great effect on horsepower requirement as the pressure ratio changes. The equations listed above have been shown in order of increasing complexity, with each calculation building on the result of the previous. From this, we can see that each new equation requires the result of the previous calculation to be accurate. This also suggests an approach when troubleshooting any control system or calculation. To begin, the pressure ratio requires accurate pressure inputs. If this is accurate, volumetric efficiency can be calculated, if we know the right clearance and slippage factors. With an accurate volumetric efficiency and piston displacement, capacity can be calculated. Finally, with capacity, horsepower can be calculated if the compressor efficiency is known. Verifying each of these in turn will allow a simplified procedure to determine the cause of inaccurate calculations.
Calculated limitations A number of compressor limitations can be calculated to help avoid mechanical problems. Some of these are: Maximum Discharge Temperature: The moving components of compressor cylinders must survive for years of operation with high pressures and loading. This reliability may be lost if components are operated above their temperature range. The valves, rings and packings of most compressors are rated for 250 degrees or higher. In some cases, components may be rated up to 350 degrees by changing to high temperature materials. The discharge temperature rise is usually calculated to assure that components are within their limits. Low Volumetric Efficiencies: Typically, controls are calculated based on an overall volumetric efficiency, where the total amount of clearance is divided by the total machine displacement. Individual cylinder ends may have considerably more clearance than the machine average, and can have low or zero volumetric efficiencies while the average value is reasonably high. For this reason, a minimum volumetric efficiency should be calculated, based on the cylinder ends with the highest amount of clearance. This can help avoid high cylinder temperatures and possible equipment failures. Maximum Pressure Differential Limited by Rod Loading: The compressor rod load is calculated based on the cylinder suction and discharge pressures and the diameters of the piston and rod. This limit must not be exceeded, or equipment damage can be expected. Excessively High or Low Horsepower: In all cases, the compressor’s driver will have a limit on its maximum horsepower. This is used as a control point to drive changes in operating condition. These may be changes in clearance for torque control, or changes in pressure or speed to restore an acceptable load. In the case of gas engine drivers, conditions of extremely low load may be damaging also. This is due to limitations of the engine, such as heat requirements to drive a turbocharger or carboning of valves and ports due to low exhaust temperatures.
11 Compressor Sizing and Application Whenever a new compressor is installed, or conditions change for existing equipment, the compressor application and sizing should be reviewed. In selecting equipment for an application, there are a number of requirements to be considered. Some of these are: 11.1 Volume requirements and cylinder size Compressors are installed to meet some specified volume requirement, usually at given suction and discharge conditions. In actual operation, pressures will range both higher and lower than the specified condition. If the compressor cylinders are designed only for the specified condition, the unit may perform poorly at other pressures. The best practice is to determine the required range of pressures, and flow extremes from minimum to maximum. Then, cylinders can be selected to deliver the required maximum volumes, and clearance or other means of unloading can be provided to meet the minimum conditions and to control driver load. As compressor cylinders are essentially a pressure vessel with a number of penetrations, they are designed for some maximum pressure rating, where typically increased cylinder diameter results in lower pressure ratings. 11.2 Cylinder Size and Rod Loading The extremes of pressure range and the compressor frame selected will put limits on the maximum cylinder diameter: All compressor frames will have some maximum rod loading limit. The imposed rod load increases as differential pressure (suction to discharge) increases, and also with increasing cylinder diameter. Once cylinders are selected for a given throughput, they must be checked against the maximum allowable rod loading for the frame or engine which is chosen. If the rod loading is excessive, it may be necessary to use a greater number of cylinders, with smaller diameters to provide the same displacement while staying within the rated loading. This option would increase the cost of the unit. Other possible compromises would be to reduce the allowable pressure range to avoid rod load, reducing the design volume to allow smaller cylinders, or using a heavier compressor frame with higher allowable rod loading. 11.3 Unloading Options Typically, some unloading will be necessary to control engine load, usually at higher discharge pressures. Drivers are usually sized to provide enough horsepower to deliver the design flow and pressures. Beyond this, unloading will be used to allow operation across the entire operating range.
Also, unloading may be used to control throughput volumes. Finally, good engineering practice is usually to select the largest cylinders possible within rod load limits, then add unloading to control load and throughput. This allows better loading of the engine when operating off the design point. This approach provides the best usage of installed equipment, but requires more unloading provisions. The most common unloading options are clearance volumes and deactivation. 11.4 Clearance Volumes Most compressors designed for gas service will have some provision for adding clearance. The best approach is where clearance is added to both ends of a double acting cylinder. When this is done, all cylinders usually have similar added clearance volumes. This keeps loading balanced between all cylinders, minimizing pulsation and vibration. This also keeps the highest possible volumetric efficiency on all cylinder ends. 11.5 Deactivation Where clearance volumes are not adequate, it may be necessary to deactivate cylinders or ends to reduce engine load or throughput. This is effective, but may result in shaking or pulsation problems. This is usually a last resort solution. 11.6 Active Load Controls Some compressors are equipped with active load controls, which allow gas to flow back through the suction valves. Controlling the amount of this backflow effectively changes the compressor displacement. This is particularly useful for storage stations, or those with wide fluctuation of operating pressure and flow conditions. It allows use of large compressor cylinders to load the engine at low pressure ratios, but also provides adequate unloading at high ratios. This type of system requires high speed actuators on the suction valves and a timing device for control of the valve action. While it reduces the compressor’s efficiency, in many cases it is the only effective means of controlling the engine load. These systems have been sold by Ingersoll-Rand and Hoerbiger Valve Company.
Glossary of Terms 1. Absolute Values- (p. 8) A measurement of pressures used in calculations of gas conditions taking gauge pressure added with atmospheric pressure. Most compressor calculations involving pressure or temperature are based on absolute values. These are referred to some absolute zero value. For pressure, absolute zero is complete vacuum. For temperature, absolute zero is approximately –4600 F 2. Blower- (p.3) Type of positive displacement compressor where two intermeshing elements rotate in an ellipsoidal chamber with inlet and outlet ports on opposite sides. 3. Compression Efficiency- (p. 15, 25) The Percentage of supplied energy that actually goes into raising the pressure of gas. 4. Compressor Packing – (p.6) Packing is used in a double acting cylinder to seal around the compressor rod, preventing gas leakage from the cylinder. Packing is normally a series of segmented metallic rings, assembled and held in the end of the cylinder by the packing case. 5. Compressor Rod/Piston Rod – (p.4,8) Cylindrical rod which connects the compressor piston to a crosshead, normally passing through a packing case to seal compression pressure into the cylinder. 6. Compressor Valves – (p.5) Compressor valves are high speed check valves, controlling flow of gas into the cylinder (suction valve) or out of the cylinder (discharge valve). They are designed for minimal pressure loss and maximum reliability 7. Crosshead – (p.4) A crosshead is a sliding component at the outer end of the connecting rod, which converts the eccentric motion of the connecting rod to pure linear, eliminating side forces on the compressor piston. 8. Crosshead Pin/Wrist Pin – (p.4) The wrist or crosshead pin connects the outer end of a connecting rod to either a single acting, trunk type piston (wrist pin) or to a crosshead (crosshead pin). 9. Degrees Rankine- (p. 22) A basis of temperature measurement related to an absolute zero of –459.70 F 10. Double Acting – (p.4) Compressor cylinder which has sealed chambers fitted with suction and discharge valves to allow gas compression on both sides of a piston. 11. Mechanical Efficiency- (p. 25) A factor to correct for mechanical friction in the compressor, covering bearing friction and the friction of compressor rings and packing. 12. Positive Displacement Compressor- (p. 2) A basic compressor type, which traps a volume of gas in a space whose volume is decreased, so that pressure increases. 13. Psia- (p. 9,24) Pressure expressed in absolute value, with atmospheric pressure included. 14. Psig- (p. 9) Pressure expressed as gauge value, with zero Psig being atmospheric pressure. 15. Reciprocating Piston- (p. 2) Type of positive displacement compressor, most widely used for gas service. Consists of piston in a cylinder with pressure actuated check valves to control suction and discharge flow through the cylinder.
16. Single Acting – (p.4) A Single Acting piston compresses gas on only one face, either by design or by deactivating valves on one side of cylinder 17. Screw/ Rotary- (p. 3) Type of positive displacement compressor containing compression chambers formed between two intermeshed elements similar to worm gears or screw threads. Designed for high pressure ratios, usually limited to pressures below 250 Psig, frequently uses oil injection for sealing/cooling. 18. Suction to Discharge Pressure Ratio- (p. 9, 12) Pressure ratio is calculated by dividing the absolute values for discharge pressure (Psia) by suction pressure (Psia). 19. Supercompressibility- (p. 23-24) Factors to correct for non-ideal compression and re-expansion of the gas. 20. Vane compressor- (p. 3) Positive displacement compressor consisting of a cylindrical chamber with a rotating paddle wheel drum mounted off-center in the chamber. 21. Volumetric Efficiency- (p.11) Percentage of compressor stroke where valves are open, compared to the entire stroke. This may be either suction or discharge volumetric efficiency.
