VERY HIGH PRESSURE COMPRESSORS 7.31 FIGURE 7.26 Cross-section of multi-bored plate. By the principle of superimposition of effects, the stress conditions generated by external pressure, internal pressure and axial preload can be considered sepa- rately. The holes are assumed to be of the through type and have a diameter which is constant, with the geometry of the valve section unchanged in any plane perpen- dicular to the valve body axis. Without axial stress, the calculation approach brings up the problem of an elastic body in a plane stress condition. Consequently, the problem consists of establishing the stress condition due to external and internal pressure in a plate geometrically schematized in Fig. 7.26. The plate has three axes of symmetry, 60 Њ apart, which correspond to the di- ameters through the hole centers. In this structure, the greatest stresses are on the inner edges of the holes, particularly on the points lying on the axes connecting two adjacent holes and on the axes of symmetry. The most interesting points (Fig. 7.26) are used to compare different calculations. The stress condition of this elastic body could be determined through an exact procedure, i.e., analytically, by solving the elastic problem, or through approximate procedures using: • Existing formulas for comparable geometrical bodies • The finite element method 25,26 • Strain gages on the piece boundary • Photoelastic models 7.32 CHAPTER SEVEN Solving a plane problem using the elasticity theory, 27 means finding the Airy func- tion. The stress function is complex due to the presence of several boundaries inside the plate and consequently the resolution of the equation system defining the elastic problem will also be very troublesome. An analytic solution of a similar case has been found by Kraus. 28 Evaluation of the stress distribution on the valve body can also be made using equations for thick-walled cylinders under external and internal pressure. 28 In the case of cylinders with a central hole, the formulæ are to establish the stress distri- bution in any point of the radial thickness. More complex are the equations for cylinders having eccentric holes, 28 giving circumferential stress in any point of the external and internal boundary. A further evaluation of circumferential stress can be made, (only for the points in Fig. 7.26) by utilizing existing studies on stress concentration factors in plates, whose notches are represented by holes. 29 In this case, the plate is assumed to be compressed uniformly, as in a solid cylinder, with the pressure acting on the outside. Variations in the circumferential and radial stresses on the required points referring to the center of the valve body being known, the circumferential stresses, resulting from the presence of the holes, can be determined. Furthermore, holes of different diameters require further simplifying assumptions. Strain Gage Method. A model of the plate was equipped with strain gages on external and internal surfaces to measure the trend of the circumferential stresses on the boundaries, with pressure acting inside and outside. 24 The model was bigger than the valve, to allow positioning of the strain gages on the internal surface and because of seal problems in the passage area of the connecting wires to the strain gages. The test pressure value was kept under 30 MPa (4350 psi). To minimize effects of systematic and accidental errors of the measuring instruments, the value of the microstrains undergoing measurement was increased, by adopting a light alloy model instead of steel, having a normal modulus of elasticity E ϭ 72500 MPa (10,512,500 psi) (about 1/3 that of the steel used for the valve). To eliminate uncertainties as to the elastic properties of this material, some specimens were taken from the piece the model was made from, to obtain the Young’s modulus and Poisson’s ratio for converting the microstrains into stresses. FEM Application. The calculations were made with pressure acting separately on the external and internal peripheries. It was assumed, according to the symmetry of the system, that there was no rotation in the nodes determining the diameters of the half-plate, and that displacement would occur only in the direction parallel to the circumference. The procedure used for calculation involved finite elements with triangular elements having three nodal points, with the general element having 6 degrees of freedom and a linear shape function, 24 whose trend of stresses is shown in the graphs in Fig. 7.27 in relation to pressure. The trend of circumferential stress with pressure acting on the outside is similar on hole edges. In fact, its lowest values comply with those predicted in points A2.1, A3.1, A2.2 and A3.2. The lowest value ( ␴ c /p e ϭϪ2.9) is assumed to be at point VERY HIGH PRESSURE COMPRESSORS 7.33 FIGURE 7.27 Circumferential and radial stresses on plate edge and symmetry axes. A.3.2, i.e., the internal boundary point of the hole having the smallest diameter and also related to the straight line joining the centers of two adjacent holes. The highest value ( ␴ c /p e ϭϪ1.9) is at point A4.2, i.e., at the smallest hole, toward the plate center and along a symmetry axis. Furthermore, with internal pressure, the curves of circumferential stresses on the inner edge of the holes show a similar trend, the highest value being point A3.2. The trend of circumferential and radial stresses is alike (Fig. 7.27), both in the case of external pressure and that of pressure in the holes. The sum of circumferential or radial stresses in the case of external pressure and unit internal pressure is constant and equal to Ϫ1, i.e. ( ␴ /p ϩ ␴ /p ) ϭϪ1 ce ci The foregoing can be proved analytically for thick cylinders with centered or ec- centric holes, as formulæ exist for stresses along the thickness and at the boundary respectively. In any case, if unit pressure exists inside and outside a cylinder, the stress condition is the same at any point of the thickness and the hoop and radial stresses are: ␴ /p ϭ ␴ /p ϭϪ1 cr This is the result of two different loading conditions, with external and internal pressure; the above equation can thus be obtained by the superimposition effect. These statements apply to any type of stress (hoop, radial or direct, according to the reference axes) involving multiconnected domains, regardless of boundary 7.34 CHAPTER SEVEN FIGURE 7.28 Comparison of theoretical and experi- mental results on multi-bored plate. shape, provided the internal pressure is considered on all internal profiles at the same time. Comparison of Results. In a polar-type representation (Fig. 7.28), the values are compared with different methods. The stresses due to internal pressure are brack- eted. 24 The trends of the curves determined according to the finite element method and the experimental measurements are similar, and the stress values are very near. The experimentally determined values, except for the central zone of the small hole, are slightly higher than those calculated with the finite elements. At the area of greatest concentration (points A3.1 and A3.2), the results practically coincide. The use of conventional equations led to results sufficiently in accordance with one another and generally lower than those obtained through the finite element method. This occurs especially at the point of greatest concentration when the thick cylinder formulæ are used. At the same points, according to the theory of notches, the results practically coincide with those obtained through the finite element method and experimental measurements. Knowledge of the effective stress condition, proper choice of materials and ob- taining a high degree of finite elements in the zones of greatest stress concentration makes it possible to arrive at the actual safety coefficient and thus ensure reliability against fatigue failure. VERY HIGH PRESSURE COMPRESSORS 7.35 7.3 PACKING AND CYLINDER CONSTRUCTION 7.3.1 Technical Solution for Cylinder Components Two solutions have been used for this special pressure vessel: • A hard metal liner (sintered tungsten carbide with 9 percent cobalt binder), shrink-fit into a steel cylinder, on which a piston equipped with special piston rings (Fig. 7.16) was sliding • A packing arrangement cup housing the seal rings, with a hard metal plunger (Fig. 7.8) Although the first solution was providing fairly good results, it was more affected by plant conditions, low polymers and catalyst carrier as the lubrication was ob- tained by injecting oil into the gas suction stream. The packed plunger solution is less influenced by such factors, considering that the lubricant is injected directly onto the sealing elements through holes and grooves on the packing cups. The technological development of sintering WC (11 to 13 percent Co) plungers of large size in one piece, the lower quantity of oil consumed, the excellent per- formance, and other process considerations 21 led to preferring packed plungers over liners on the compressors manufactured in the last 25 years. The selection of materials for components under pressure is very important. Mechanical properties must always be carefully analyzed and, when extreme fatigue conditions exist, aircraft-quality electroslag or vacuum arc remelted steels should be utilized. To obtain adequate fatigue strength of pressure components, it is necessary to use autofrettage when operating pressures are very high. Sealing surfaces between cylinder components play an important role in achiev- ing good cylinder performance. These are normally flat annular surfaces lapped to a finish of 0.2 microns CLA* and pressed together by tie rods so that their resulting load provides sufficient contact pressure to achieve seal. Since little can be done to modify the actions the cups are subjected to during operation, care should be taken to prevent the consequences of accidental surface defects by performing local precompression treatments, such as cold rolling, shot peening, ionitriding etc. Special attention is required for the surface finishing of elements in direct contact with the fluid subjected to pulsating pressure. In order to eliminate superficial faults as much as possible, which could cause fatigue failure, very high grade finishes are required. Tungsten carbide plungers and liners have surfaces with 0.05 microns CLA; with the additional advantage of reducing to a minimum the coefficient of friction between the moving parts. It is difficult to obtain these low roughness values on the gas passages in the cylinder heads and on the surfaces of steel cylinders in general, without the use of special machinery. *CLA ϭ Center Line Average. 7.36 CHAPTER SEVEN With a surface finish of 0.8 microns (32 microin.), fatigue life is reduced by 15% as compared to that of a finish of 0.025 microns (1 microin.) It is not necessary to obtain perfectly smooth surfaces, as it has been proved that finishes of 0.1 microns (4 microin.) have no greater fatigue resistance than surfaces with roughness of 0.025 microns (1 microin.). 7.3.2 Sliding Seals Between Piston and Cylinder at Very High Pressures The contact between the sliding parts for adequate sealing is severe for packing and particularly for piston rings. Under normal pressure, as the relative movement is parallel and does not allow perfect lubrication, only a transient condition of film lubrication and dry friction exists. Oil particles, between the contact points, prevent galling, but to keep the friction coefficient within allowable limits, and to avoid excessive heat generation, a correct selection of materials (chemical and physical properties) is necessary. Experience has shown that the most suitable materials for sealing elements are bronzes, having good wear resistance and mechanical prop- erties. Cast iron and bronze or various combinations of these metals were used in the past for piston rings. Special bronzes are still utilized for packing sealing elements, although plastic elements can be used up to 250 MPa (36250 psi) when the process requires low heat generation to avoid decomposition in the cylinder. Relating to the plunger material, in the past, nitrided steels were used for plungers in ammonia compressors up to 100 Mpa (14500 psi). Usually, today piston rods are made of steel coated with tungsten carbide (11 to 13% Co) up to pressure of 60 MPa (8700 psi). In polyethylene plants, with more severe pressure conditions and more precari- ous lubrication by white oils, liners or plungers are made of tungsten carbide with cobalt bonding. When the cobalt content is increased, the hardness decreases, but the toughness increases, and this quality is more important for plungers than for liners. Today, the steel plunger coated with tungsten carbide can be used up to 140 MPa (20300 psi), usually on the first stage of secondary compressors. The sliding surface of plungers and liners should be machined to the maximum degree of finish obtainable in order to reduce the friction coefficient to a minimum. Values of 0.025 to 0.05 microns (1 to 2 microin.) CLA of roughness are normally achieved. In case of WC coated plungers, the surface roughness is 0.1 microns CLA (4 microin.). The surfaces of sealing elements do not require the same high quality, since they are softer and on the plunger they are polished during operation, but still need lapped mating surfaces and more accurate geometry to prevent leaks and failures. The life of the sealing elements is influenced by other factors. The stroke and revolutions per minute (RPM) determining the average piston speed influence the life, since heat generation increases with speed. The RPM are limited by compres- sor size and arrangement, dynamic loads on the foundation, operation of the cyl- inder valves, and pressure pulsation in the gas pipes. VERY HIGH PRESSURE COMPRESSORS 7.37 The stroke is selected to have a mean piston speed between 2.7 and 3.3 m/s (530 to 650 ft/min). A long stroke is generally desirable since this exposes a longer part of the plunger out of the packing, for more effective cooling. The life of sealing elements is influenced by the system supplying the oil to the cylinder, the amount and quality of oil, the shape of sealing elements, and the linearity of plunger movement. A continuous film of oil must be applied to the sliding surfaces. The type of oil is selected mainly for process reasons (i.e., the need to keep the product pure), and also its lubricating properties. It is current practice to use white oil. The shape of the sealing elements used is similar to those used in conventional machines. The piston rings solution, with lubricating oil entrained by the gas, needs only few rings for efficient sealing, but also to enable the one most distant from the gas- oil mixture to be lubricated. Each combined piston ring is made of two rings in the same groove, with a further ring mounted beneath. The ring gaps are positioned out of alignment to give a complete seal effect. On the top of the rings there is a bronze insert, improving the anti-friction properties and the running-in. A packing arrangement is usually composed of 5 elements, for pressures up to 350 MPa (50750 psi). In the past, solutions with 3 to 8 sealing elements were also applied. The ring nearest the pressure is a breaker ring of special shape, suitable for damping the high pressure fluctuations but not designed to provide effective seal, as this function is performed by the following ring couples, whose life is consequently increased. The amount of oil applied must be controlled accurately, since trouble can arise from either excessive or insufficient lubrication. If excessive oil is injected and the seal rings are providing perfect seal, the oil pressure can rise to a value above that of normal conditions and the contact pressure between rings and plunger could cause seizure. Of great importance is the linearity of the piston movement, since it ensures that the sealing elements will not be subjected to irregular operating conditions and thus forced to assume an incorrect position in their housing, with consequent overstressing and reduction in life. It is necessary to keep the temperature low by cooling the plunger with oil around it, outside of the main packing. This is important mainly to reduce the risk of thermal cracks on the plunger surface. 7.3.3 Autofrettage of Various Cylinder Components General Aspects. The use of autofrettage, applied to tubular and vessel-reactors, has been extended to pumps 18 and to machines operating particularly in tubular- reactor plants, as it is effective where the probability of fatigue failure is high. This technique allows components to be built using materials with lower mechanical properties. Autofrettage is performed on cylinder heads with combined axial valves, when high pressures are involved, as gas pulsations are still present and fatigue must always be taken into consideration. Cylinder chambers and packing cups are ex- cluded, as they can reach adequate prestress levels through shrink-fitting. Packing 7.38 CHAPTER SEVEN FIGURE 7.29 Surface seal with conical seat. cups with axial holes and oil distribution cups require additional prestress only inside the lube oil hole. The distribution cup has no shrinkage, and generally has curved holes normally obtained through a special procedure, such as electro dis- charge machining (EDM). In this case, a proper polishing procedure should be applied to fully remove the surface modified by local defects. Autofrettage of injection quills and check valves operating on ethylene second- ary compressor second stages is also common practice when pressures are very high. Cylinder heads with radial valves are shrink-fit and are autofrettaged only when differential pressure between suction and discharge is very high. Autofrettage pressure is determined by operating conditions, geometry, presence of prestresses (due to shrink-fitting), and properties of the material. Autofrettage pressures for hypercompressor cylinder parts range between 500 MPa and 1300 MPa (72500 to 188500 psi). 30 Autofrettage of axial holes is performed after shrink-fitting of the cup on the finished piece, only upon completion of machining before final lapping of the mating surfaces. In this case, autofrettage pressure has been applied up to 1100 to 1300 Mpa (160000 to 188500 psi). Test Rigs and Seals Arrangement. Few types of seals withstand very high pres- sure applications, due to the fact that the geometries of the cylinder components to be autofrettaged are often complex. On polyethylene compressor cylinder parts, seals are restricted to conical seating surfaces, metal gaskets, plastic O-rings and special arrangements: 30 • The cone solution (Fig. 7.29), typical of high pressure tubing, has been applied up to 1300 MPa (188500 psi). • Annealed copper gaskets are used up to 1300 MPa (188500 psi) (Fig. 7.30). • Viton O-rings are employed for small-diameter seats, tapered (Fig. 7.31) or flat (Fig. 7.32), protected against extrusion by the metallic contact between the parts. VERY HIGH PRESSURE COMPRESSORS 7.39 FIGURE 7.30 Metal seal. FIGURE 7.31 Plastic O-ring with conical seat. Positive results were obtained on diameters up to 76 mm. (3 in.) and up to 900 MPa (130500 psi) for the latter solution. • Self-sealing arrangements (Fig. 7.33) are used for wider diameters, in order to follow the bore, subject to considerable strain under high pressures. These seals are made as follows: • A seamless plastic O-ring with hardness between 75 to 90 Shore A, with good surface finish • Hard plastic (a polyamide resin) and geometrically precise shoulder rings. Di- mensions have to be carefully checked, as plastics are subject to alteration with the passage of time. • Bronze antiextrusion rings with a 45Њ angle • Bronze rings to preload the seal assembly and to guide the inner core of the device In autofrettage of radial valve cylinder heads, similar seals are used and internal mandrels are applied to reduce fluid volume. Axial valve cylinder heads are auto- frettaged (Fig. 7.34) with special seals (Fig. 7.33) to achieve seal on the large inner diameter which can be accomplished by providing a smooth surface finish and 7.40 CHAPTER SEVEN FIGURE 7.32 Plastic O-ring with flat seat. FIGURE 7.33 Special seal with O-ring. using great care in assembling the rig to avoid local damage in the seal zone. An internal bar reduces fluid volume. The seals are preloaded, the assembly is balanced and no additional support is required for the inner core. Lateral (suction and dis- charge) holes are plugged by flanges using combined metallic and O-Ring seals (Fig. 7.32). Autofrettaged packing cup axial holes (Fig. 7.35) use metal seals (Fig. 7.30). The test rig for the oil distribution cups uses axially-directed seal (Fig. 7.30) and radial seal (Fig. 7.31). Autofrettage of injection quills utilizes cone seals (Figs. 7.29 and 7.31). Autofrettage Procedure. In equipments operating at very high hydrostatic pres- sures, the fluid must be able to transmit pressure without undergoing freezing ef- fects, related to fluid properties, operating temperatures and tubing size. Pressure may increase at the pump and, due to solidification problems within the tubing, may be much lower inside the piece to be autofrettaged. Brake oils have been used up to 500 MPa (72500 psi) with some drawbacks (i.e., corrosion on pump seal rings caused poor performance). Prexol 201 over- comes solidification problems and gives adequate intensifier plunger seal life, up [...]... standards tighten, consideration must be given to CNG compressor manufacturer’s gas leakage rate data TABLE 8.1 Compressor Stages vs Suction Pressure Suction pressure (psig) 6؆ H2O-10 6؆ H2O-100 80– 350 250 –1200 1000ϩ No of stages Discharge pressure (psig) 5 4 3 2 1 3600 50 00 3600 50 00 3600 50 00 3600 50 00 3600 50 00 8.4 CHAPTER EIGHT Pressurized Crankcase Compressors with pressurized crankcases collect seal... density The compressor is used to boost the pressure of natural gas and is the primary equipment of the compressed natural gas (CNG) refueling station 8.2 CNG COMPRESSOR DESIGN The compressor type used is the multi-stage reciprocating piston compressor Compressor size commonly ranges from 25 to 250 brake horsepower (BHP) The design of the CNG compressor resembles the high pressure air compressor but... Bernardini, ‘‘Aspects of Research on Secondary Compressors for Low Density Polyethylene Plants,’’ Quaderni Pignone 25, June 1978, pp 123–124 5 Vinciguerra, C., U.S Patent 3, 58 1 .58 3 to Nuovo Pignone S.p.A., January 15, 1969 6 Andrenelli, A., ‘‘Special Features in Reciprocating Compressors for Polyethylene Production,’’ Proceedings of the Industrial Reciprocating and Rotary Compressors: Design and Operational... Behaviour of High Pressure Packing Used in Secondary Compressors for Low Density Polyethylene Production,’’ Proceedings of the 2nd Int Conf on H.P Engineering, University of Sussex, Brighton, England, July 8–10, 19 75, pp 57 58 24 Giacomelli, E., ‘‘Finite Element Method on Polyethylene Compressor Valves Design,’’ Quaderni Pignone 26, January 1979, pp 19– 25 25 Zienkiewicz, O C., ‘‘Axi-Symmetric Stress Analysis,’’... Similar to air compressors, the natural gas compressor must be depressurized for start up This necessitates that on shutdown, gas entrapped in the compressor and piping system must be vented Unlike an air compressor which can be vented to atmosphere, the natural gas compressor must be provided with a blow down gas receiver tank This tank must be adequately sized 8.6 CHAPTER EIGHT to allow the compressor. .. commonly specified noise level is 75 dba measured at 10 feet from the perimeter of the compressor skid Most compressor packagers can meet this noise level with an enclosed and sound attenuated compressor skid package 8.2.6 Compressor Electrical Systems Natural gas, being flammable, requires that all electrical equipment and wiring within a code specified distance from natural gas compressors and gas containing... equivalent per day 8 .5. 1 Compressor Selection Table 8.4 indicates the number of NGVs that will fill up each business hour Total amount of gas dispensed each day will be 75 NGVs ϫ 4.6 gal equiv ϫ 108.7 scf/gal equiv ϭ 37 ,50 0 scf This is also the amount of fuel that must be compressed daily A maximum number of compressor operating hours per day is arbitrarily set at 8 Fewer operating hours increase compressor size,... operation More operating hours reduce compressor size and the ability to meet random increases in refueling demand The required compressor flow capacity is calculated 37 ,50 0 scf/8 hrs/60 min/hr ϭ 78 scfm From Fig 8.2, a 25 psig suction pressure and 3600 psig discharge pressure requires 0.48 BHP/scfm At 78 scfm, the required compressor horsepower is 37.4 BHP A 40 BHP, 4 stage compressor is selected, providing... maximum of 0 .5 lb/mmscf at compressor discharge has become a common industry standard This standard can be met using nonlubricated compressors or lubricated compressors with filtration In deciding lubricated versus nonlubricated, other factors to consider are outlined in Table 8.2 8.2.4 Piston Ring and Seal Performance The extreme gas pressures exerted in the final stages of a natural gas compressor present... a pressure rise of 10 MPa (1 450 psi) per second is on the safe side Generally, the test requires a pressure rise of 5 minutes minimum Pressure increase is related to the volume of the fluid in the whole system and its components (tubing and the intensifier) The autofrettage pressure is maintained for 15 minutes (5 minimum) and a slow pressure decrease takes place in about 5 minutes Slow return to final . Secondary Compressors for Low Density Polyethylene Plants,’’ Quaderni Pignone 25, June 1978, pp. 123–124. 5. Vinciguerra, C., U.S. Patent 3, 58 1 .58 3 to Nuovo Pignone S.p.A., January 15, 1969. 6 COMPRESSOR DESIGN The compressor type used is the multi-stage reciprocating piston compressor. Com- pressor size commonly ranges from 25 to 250 brake horsepower (BHP). The design of the CNG compressor. rings varies from 55 to 80 Brinell (measured with a 10 mm. ball and 50 0 kg. load). The plunger on which the sealing elements slide is made of solid tungsten car- bide, with surface finish of 0. 05 microns