COMPRESSOR HANDBOOK phần 7 ppsx

RECIPROCATING COMPRESSOR SEALING 17.21 FIGURE 17.15 Static seal arrangement in a packing case. regulatory limit for compressor leakage. It is indirectly established by the limit on the allowable concentration of VOC (volatile organic compounds) measured near the leak source. Federal regulations consider 500 ppm (parts per million) to be ‘‘no detectable emissions.’’ The regulations require that any compressor not in compliance with that emission level must have a barrier fluid (gas or liquid) system installed on its seal. A ‘‘barrier’’ system is one in which a non-VOC liquid, or gas, is forced to flow into the seal in a direction opposite to that of the leakage. The barrier fluid or gas blocks the escape of leakage to atmosphere from the compressor seal or packing. The system can be installed in one of two ways: 1. A fluid, usually inert, is injected to form a barrier seal between the process gas, and atmosphere. The fluid pressure must be maintained slightly above the pres- sure upstream of the seal. 2. More commonly, a barrier gas is applied between a set of ‘‘barrier seals’’, usually WAT or AL rings as illustrated in Fig. 17.16. The barrier gas is held at a pressure exceeding the pressure in the vent which carries leakage away from the compressor seal. On reciprocating equipment the motion of the rod will carry fluid, usually oil, in the form of a thin film under the seal face and out to atmosphere where emissions of gas carried in the oil may be released. The oil film on the rod surface absorbs gas while in the cylinder due to the relatively high gas pressure, and then releases this gas to atmosphere when the rod moves out of the packing. On non-lube ap- 17.22 CHAPTER SEVENTEEN FIGURE 17.16 Typical purged packing. plications, gas molecules can be carried out of the packing in the rod surface irregularities. In either case, a small quantity of gas may escape the barrier seal and become fugitive emissions outside the packing flange. These may measure as much as 200 ppm, but generally are much lower. Gas transported by the rod surface becomes important only when allowable emissions approach zero. 17.21 WIPER PACKING In addition to wiping oil from the rod surface and returning it through a drain back to the crankcase, wiper packing also has a sealing function. Pressure pulses from the crosshead, acting as a piston, may cause a ‘‘breathing’’ action through the wiper that will cause it to leak oil over into the distance piece. There can also be leakage of gas from the distance piece into the crankcase. The seal rings, as indicated in Fig. 17.17, must minimize both these leakages. The seal and wiping function can be combined into one groove if it is necessary to seal only the crosshead pressure pulse. This single groove would normally con- tain a butt tangent cut ring paired with two wiper rings. Compared to the three wiper ring combination, this ring sacrifices some ability to scrape oil effectively from the rod, but it is sufficient for slow speed compressors. To wipe oil from the rod surface requires an apparent contact pressure between ring and rod of about 50 psi (controlled by the garter spring). Lower pressure tends to leave a thicker film, while high pressure may allow rapid wear of ring edge or rod surface. Ring to rod fit and contact is the most important factor in wiper packing performance. RECIPROCATING COMPRESSOR SEALING 17.23 FIGURE 17.17 Typical wiper packing arrangement. From a wiping standpoint, metal or hard plastic works best for the wiper ring. However, seal rings, if made of a softer, more conformable plastic (typically filled TFE), offer the best performance from a wear, friction, and sealing standpoint. The wiper packing case should have a drain(s) shielded from oil spray or splash that results from crosshead motion. The force of this oil thrown up by the crosshead can, in some instances, block free drainage away from the wiper rings. Establishing good open drains as well as proper ring/rod contact becomes more important on the smaller, higher speed, compressors. 17.22 HIGH PRESSURE (HYPER) PACKINGS ‘‘Hyper’’ is generally taken to mean over 10,000 psi. (Compressor discharge pres- sure might go above 100,000 psi.) At these pressure levels, fluids are usually acting very much like liquids in that compressibility is low. The type of rings used to seal these pressures are similar to those used in lower pressure compressors, but ring designs and materials must be selected to withstand high ‘‘compressive’’ cyclic pressures. The problems (wear, friction and heat) caused by high contact pressure between ring and plunger are normally overcome by using ring sets that act partly as a labyrinth, and thus spread the pressure drop across several rings. Life of hyper packings is increased also by high lubrication rates plus, in some cases, coolant (oil) flow across the rod downstream of the packing. The other principle problem in hyper packings is the containing of very high cyclic pressure within the case, which is essentially a thick-walled pressure vessel. Particular attention has to be paid to stress concentrations, such as holes or notches, 17.24 CHAPTER SEVENTEEN FIGURE 17.18 High pressure packing with compounded packing cups and oil circulating around plunger and case. FIGURE 17.19 High pressure packing with discharge pressure surrounding the case. that might raise stress beyond acceptable levels. Compounding of cups, autofret- taging, or pressure loading the outside of the case are typical ways to ensure long life of the case parts. Typical high pressure packing is shown in Figs. 17.18 and 17.19. 17.23 COMPRESSOR PISTON RINGS Although sealing principles for rings on the piston are the same as those in the packing, their construction is somewhat different. In normal compressors, the re- quirement for sealing at the piston is not as stringent as it is in the packing. In fact, there is some reduction in wear if piston rings do leak slightly and thus RECIPROCATING COMPRESSOR SEALING 17.25 FIGURE 17.20 Piston ring types. distribute the pressure drop over more than one ring. The predominant type of joint for double-acting cylinders is the angle, or butt cut as in Fig. 17.20. For single- acting cylinders where leakage is more important, a seal joint is sometimes used. Regardless of the joint style, a large percentage of compressors use segmental rings—either two- or three-piece. The segmental type allows a ring with more radial thickness, which exerts less load against the cylinder wall than a radially thin one-piece ring. Choosing the number of piston rings to use is, to some degree, an art. The quantity, no doubt, influences one thing of primary importance—ring life, and attempts have been and are being made to put this in to a good relation. However, at the moment, the number of rings selected for most applications is based essen- tially on experience. A guide for number of sealing rings generally used is included with Fig. 17.21. 17.24 COMPRESSOR RIDER RINGS In some lubricated applications, and all nonlube ones, it is necessary to use a separate bearing, or rider ring(s), on the piston. This can be metal or plastic and serves only to keep the piston from contacting the cylinder. The rider must be made wide enough to keep bearing pressure between rider and cylinder very light, since 17.26 CHAPTER SEVENTEEN FIGURE 17.21 Typical piston ring arrangements and number of rings required versus pressure dif- ferential. tolerance for wear will be less than for piston rings. This is because even with a rider, relatively little clearance separates the piston and the cylinder, and no more than this can be worn from the rider before piston and cylinder come into contact. One major problem with riders is preventing them from pressure actuating like the sealing rings. They are usually notched on the sides or across the face and, in some instances, grooved or drilled in such a manner that they will not trap gas and thus seal like piston rings. They can be made either with a cut, as shown in the illustration, or uncut. Both of these types have advantages and disadvantages. The uncut ring is more difficult to install, will not tolerate even moderate temperature increases, but is slightly less prone to act as a seal as long as it remains tight against the groove bottom. A cut ring is easily installed, has room for expansion circumferentially, and has the advantage of large end clearance, through which gas can readily flow. The rider supports piston weight plus one-half rod weight. This load is consid- ered to be carried by the projected contact area of a 120 Њ arc. Loading is usually acceptable if kept below 5 psi for nonlubricated cylinders. For lubricated service, American Petroleum Institute Standard 618 limits rider loading to 10 psi, but this has been extended to 50 psi successfully in a number of applications. 17.25 PISTON RING LEAKAGE The average compressor has from two to six piston rings. The most common ring joint is an ‘‘open type’’ butt, or angle cut as in Fig. 17.20. Leakage wise, these are RECIPROCATING COMPRESSOR SEALING 17.27 FIGURE 17.22 Instantaneous pressure between piston rings. about the same. There is some slight advantage to the angle cut, but this is often overshadowed by the other factors affecting leakage. Nearly all the leakage occurs through the joint since this is the only point in the ring where there is a definite opening or orifice. The opening is a rectangular passage with one dimension equal to the ring gap and the other to the piston clearance. This path is subject to wide variation—it will be almost zero when positioned at bottom of the piston, changing to maximum when at the top. It also constantly increases as the ring wears. Both of the leak path dimensions are a function of cylinder diameter, so in general, leakage can also be related to cylinder diameter. Pressure distribution across the rings has been analyzed and measured and, for a two-ring piston, would look as illustrated in Fig. 17.22. For additional rings, this becomes more complicated, but in essence pressure within the ring pack cycles through a range somewhere between suction and discharge, resulting in a differ- ential across the rings, first in one direction and then in the other. Leakage through the rings then is not a result of steady pressure drop, but changes constantly. Leakage can be expressed however, as an average flow of gas during any particular compression cycle. A representation of approximate leak rates is pictured in Fig. 17.23. Quantity wise, there can be large variations. For example, .03 scfm in a small cylinder all the way up to 40 in a very large one. In a double-acting cylinder this is not actually leakage, as gas is not lost. It simply passes from one side of the piston to the other. It is really a loss only from compressor discharge capacity. So, a better way to look at this is as a percentage change leakage causes to cylinder volumetric efficiency. For new rings in lubricated applications, loss of V.E. with open joint rings will be about .5% up to approxi- mately 3%. 17.28 CHAPTER SEVENTEEN FIGURE 17.23 Average leakage by pis- ton rings (for light gases in a double acting cylinder at ratio of 4). Rings in nonlube cylinders suffer in two respects. First, no oil is present to reduce leak paths, and second, piston clearances are larger. Both of these have the effect of increasing leakage by roughly three times. To recover the loss of capacity, seal joint rings are often used as in Fig. 17.20. These rings have no theoretical leakage paths, plus leakage remains, essentially constant during the life of the ring, that is, until the gaps open. Based on tests, it is a good assumption these styles of rings will reduce leakage, when compared to open gap rings, by as much as 90%. 17.26 COMPRESSOR RING MATERIALS The trend to plastics has not completely left metals behind. For lubricated service, time-proven bronze and cast iron are still commonly used materials. These are good simply because they are excellent bearing materials. They have the ability to carry and hold lubricant because of their porosity, the chemistry or structure to supply their own lubrication when oil is lacking, and heat transfer properties to quickly carry frictional heat away from the rubbing surface. To replace these metals with plastics with equally good properties requires se- lection from an almost infinite number of plastic-filler or plastic-plastic composites. The first plastics that made successful rings were the phenolic and cloth laminates. These are resistant to many chemicals, will work under marginal lubrication, and are relatively inexpensive. The other important group has been the low-friction, but weaker, plastics blended with a strengthening filler or another stronger plastic. In this last group, there have been only a few with frictional properties good enough to run without lubrication. RECIPROCATING COMPRESSOR SEALING 17.29 It is difficult to put plastics in categories, but the most useful as related to compressors might be described as thus: *Higher friction plastics needing at least some lubrication: Polyimide (PI) Poly(amide-imide) (PAI) Polyetheretherketone (PEEK) Polyphynylene Sulfide (PPS) Polyamide (Nylon) Phenolic-Cloth laminates Low-friction materials capable of running without lubrication: TFE plus strengthening and wear reducing fillers PI plus friction reducing fillers PEEK with friction reducing fillers This is just a general grouping of the materials, but indications are that from these come most of the best, or most common, seal ring materials. The properties of these materials influence the types of piston and packing rings used. For example, the relatively low yield strength of TFE blends has dictated the use of Type TR rings; the high elongation and ‘‘plastic memory’’ of TFE allows its use in stretch-on riders, while the strength and stiffness of some newer plastics make them useful for BT or M rings or as anti-extrusion rings. This, plus the fact that compressors face such a variety of conditions, is the reason there may never be universal ring ‘‘standards’’ in material or configuration. As new materials come along, the rings applied to compressors are designed around material properties, as well as operating conditions. 17.27 SEAL RING FRICTION More than any other characteristic, friction serves as an indicator of how well compressor seals are performing. Like wear, this is very dependent upon lubrica- tion. A certain amount of power must be put into a compressor to overcome seal ring friction, but as indicated in Fig. 17.24, this is relatively low compared to the power needed for gas compression. These curves are based on normal size piston and rod packing rings. Power needed to overcome ring friction will usually be only about .5% to 2% of the compression HP. Essentially, all the frictional horsepower changes to heat * All these materials have friction reducing fillers. 17.30 CHAPTER SEVENTEEN FIGURE 17.24 Power required (heat generated) to overcome ring friction versus ring diameter. and, if not conducted away from the packing or piston rings, can cause wear and leakage. To illustrate the affect of this heat, the approximate 1.6 HP (68 BTU/ Min.) as shown, which might be generated on a 3 inch rod, will cause approxi- mately 100 ЊF rise in four minutes if not conducted away. The principle ways to reduce friction are proper lubrication, low friction materials, and narrow rings. 17.28 COOLING RECIPROCATING COMPRESSOR PACKING One of the critical, if not most critical, factors in obtaining good service from compressor packing is proper cooling. A primary source of heat is from the work required to overcome frictional resistance of the seal rings. This is influenced by material selection, ring dimensions, characteristics of the compressor, and operating conditions. The relation between cooling requirements and the various influencing factors is not known precisely. What follows is intended to serve as a guide, indicating when special cooling is required and to help in sizing the equipment needed to provide the cooling. Low friction materials, such as TFE blends, or carbon graphite, have made it possible for packing and piston rings to operate without lubrication. Frictional characteristics of these materials are good, but not nearly as good as when lubri- cation is used. Configuration of the seal rings affects this somewhat, but with most designs considerable frictional heat is generated. The primary purpose of cooling packing is to remove heat generated due to friction between seal rings and the rod. Nearly all the work done to overcome friction converts to heat at the ring and rod mating surface. This heat is transferred to the case, gas passing through the cylinder, distance piece, and the crankcase. [...]... maintain a reasonable seal RECIPROCATING COMPRESSOR SEALING FIGURE 17. 26 17. 33 Friction loading for various packing ring types FIGURE 17. 27 Coefficient of friction for various materials and levels of lubrication The value for ƒ, along with ring dimensions, operating pressures and rod speed can be used to calculate BTU/Min generated by the seal rings (From Eq 17. 2).) 17. 28.2 Heat Transfer to Gas The heat... EIGHTEEN FIGURE 18.4 COMPRESSOR LUBRICATION FIGURE 18.5 18.8 18.9 Compressor seal life vs oil usage NON-LUBE (NL) COMPRESSORS There are certain compressor applications where no oil can be tolerated in the gas stream To be most effective, these compressors require PTFE piston rings, packing rings, wear bands, and nonmetallic compressor valve parts, plus a discharge gas temperature under 275 ЊF They also... coefficient of friction is independent of load and contact area The work required to overcome friction and the heat generated is: Work ϭ (F)(2)(S) Revolution ( 17. 1) (F)(FPM) 77 7 ( 17. 2) BTU/MIN ϭ where: F ϭ Force required to move rod against friction (lb.) S ϭ Compressor stroke length (In.) FPM ϭ Average rod speed (Ft/Min.) Velocity of the rod is not constant throughout the stroke, but again, the coefficient of... the less expensive plastics do not have the frictional properties to allow them to be effective for nonlube service  17. 36 CHAPTER SEVENTEEN FIGURE 17. 28 Examples of coolant calculations Between these two extremes, compressors operate in mini- or semi-lube service or, as shown in Fig 17. 27, ‘‘poorly lubricated.’’ In this type of service, it is more difficult to select optimum material to provide the lowest... FIGURE 18 .7 COMPRESSOR LUBRICATION 18.13 other lubricator system, it must have a no-flow shut-down device to prevent compressor operation when no oil is flowing Reliable operation can only be achieved when ‘‘balancing valves’’ (Fig 18 .7) are in each divider block discharge line and manually adjusted for higher than the highest cylinder pressure 18.10.3 Cylinder Check Valves Each oil injection point on compressor. .. Approximate range of application of various compressors FIGURE 19.2 Operating conditions of centrifugal compressors.6 19.5 19.6 CHAPTER NINETEEN FIGURE 19.3 Schematic of bearing locations in reciprocating compressors The crosshead usually has a bronze bushing operating in the boundary lubrication regime 19.2.2 Bearings in Centrifugal Compressors By far the most commonly used compressors are of the centrifugal... of reciprocating compressors, the crankcase oil lubricates the bearings, carries away bearing heat, reduces friction, and prevents corrosion In reciprocating compressor cylinders, the oil is a once through operation designed to reduce friction and wear as well as preventing corrosion 18.1 ROTARY SCREW COMPRESSORS Figure 18.1 shows a typical piping flow diagram for an oil-flooded screw compressor Oil is... cavitation region for those oils on a cold start 18.3 COMPRESSOR CYLINDERS Compressor cylinder lubrication is completely different in that oil passes ‘‘once through’’ the cylinder with no recycling Successful operation depends upon uninterrupted, continuous, metered flow to a cylinder bore and piston rod COMPRESSOR LUBRICATION FIGURE 18.2 18.3 Oil flow diagram compressor crankcase The gases being pumped range... adequate for compressor crankcase oil operating at 140 ЊF, but totally inadequate for air compressor cylinders operating over 300 ЊF As little as 2% excess 02 in a gas stream will cause serious ash type build up in a matter of weeks In order to handle this problem, the compressor oil must be fortified with a high-temperature, anti-oxidation inhibitor 18.5.4 Anti-Foam Additive Lube oil leaving a compressor. .. is Eq ( 17. 4), for calculating the surface coefficient for gas passing over a smooth surface hc ϭ 05 k where: hc k V u d ϭ ϭ ϭ ϭ ϭ ͩͪ 75 Vd u ( 17. 4) Film coefficient (BTU/Hr-Ft2-ЊF) Thermal conductivity (BTU-Ft/Hr-Ft2-ЊF) Gas velocity (Ft/Hr) Gas viscosity (Lb/Hr-Ft) Gas density (Lb/Ft3) The total heat flow, Q, into the gas, may be expressed as Qϭ where: Q ⌬T D S ϭ ϭ ϭ ϭ (hc)(⌬T)(D)(S) 5500 ( 17. 5) Heat . the heat generated is: Work ϭ (F)(2)(S) ( 17. 1) Revolution (F)(FPM) BTU/MIN ϭ ( 17. 2) 77 7 where: F ϭ Force required to move rod against friction (lb.) S ϭ Compressor stroke length (In.) FPM ϭ Average. 17. 18 and 17. 19. 17. 23 COMPRESSOR PISTON RINGS Although sealing principles for rings on the piston are the same as those in the packing, their construction is somewhat different. In normal compressors,. or to maintain a reasonable seal. RECIPROCATING COMPRESSOR SEALING 17. 33 FIGURE 17. 26 Friction loading for various packing ring types. FIGURE 17. 27 Coefficient of friction for various materials