ARNOLD, K. (1999). Design of Gas-Handling Systems and Facilities (2nd ed.) Episode 2 Part 4 pdf

Reciprocating Compressors 311 Single-acting cylinder, crank end Double-acting cylinder where RL C = rod load in compression, ib RL, = rod load in tension, Ib ap = cross-sectional area of piston, in. 2 P d = discharge pressure, psia P s = suction pressure, psia P u = pressure in unloaded area, psia a r = cross-sectional area of rod, in, 2 The calculations shown above provide the gas load imposed on the rod (and crosshead bushing) by the compressor cylinder piston. To provide a reasonable crosshead pin bushing life, the rod loading at the crosshead bushing must change from compression to tension during each revolu- tion. This is commonly referred to as "rod reversal" and allows oil to lubricate and cool one side of the bushing while load is being applied to the other side of the bushing. A single-acting, head end cylinder will not have load reversal if suc- tion pressure is applied to the crank end. Similarly, if discharge pressure is applied to the head end of a single-acting, crank end cylinder, load reversal will not occur. In addition to the gas load, the rod and crosshead pin bushing is sub- ject to the inertia forces created by the acceleration and deceleration of the compressor reciprocating mass. The inertia load is a direct function of crank radius, the reciprocating weight, and speed squared. The total load imposed on the crosshead pin and bushing is the sum of the gas load and the inertia load and is referred to as the "combined rod load." The combined rod load should be checked anytime the gas loads are approaching the maximum rating of the compressor frame or anytime rod reversal is marginal or questionable. 312 Design of GAS-HANDLING Systems and Facilities COOLING AND LUBRICATION SYSTEMS Compressor Cylinder Cooling Traditional compressor cylinder designs require cooling water jackets to promote uniform distribution of heat created by gas compression and friction. Some of the perceived advantages of water-cooled cylinders are reduced suction gas preheat, better cylinder lubrication, prolonged parts life, and reduced maintenance. Operating experience during the last 30 years has proven that com- pressor cylinders designed without cooling water jackts (non-cooled) can successfully operate in most natural gas compession applications. Some of the perceived advantgaes of non-cooled cylinders are simplified cylin- der designs that reduce cost and improve efficiency, reduced initial sys- tem costs due to reductions in the cooling water system, improved valve accessibility, and reduced weight. Many manufacturers, users, and compressor applications still require that compressor cylinders be supplied with liquid-cooled cylinders. Fig- ure 11-21 includes schematics of several types of liquid coolant systems. In static systems, the cooling jackets are normally filled with a glycol and water mixture to provide for uniform heat distribution within the cylinder. This system may be used where the AT of the gas is less than 150°F and discharge gas temperature is less than 190°F. Thermal siphons use the density differences between the hot and the cold coolants to establish flow. This system may be used where the AT of the gas is less than 150°F and discharge gas temperature is less than 210°F. Forced coolant systems using a mixture of glycol and water are the most common for natural gas compressors. Normally, the compressor cylinder cooling system and compressor frame lube oil cooling system is combined. A single pump is used to circulate the coolant through the cylinders and the lube oil heat exchanger and then to an aerial cooler where the heat is dissipated. When forced coolant systems are used, care must be taken to provide the coolant at the proper temperature. If the cylinder is too cool, liquids could condense from the suction gas stream. Thus, it is desirable to keep the coolant temperature 10°F higher than that of the suction gas. If the cylinder is too hot, gas throughput capacity is lost due to the gas heating and expanding. Therefore, it is desirable to limit the coolant temperature to less than 30°F above that of the suction gas. Reciprocating Compressors 313 Figure 11-21. Cylinder cooling systems. (Reprinted with permission from API, Sid. 618, 3rd id., Feb. 1986.) Frame Lubrication System The frame lubrication system circulates oil to the frame bearings, con- necting rod bearings, crosshead shoes, and can also supply oil to the packing and cylinder lubrication system. Splash lubrication systems are 314 Design of GAS-HANDLING Systems and Facilities the least expensive and are used in small air compressors. Forced-feed systems are used for almost all oilfield gas compression applications. Figure 11-22 shows a splash lubrication system where an oil ring rides loosely and freely on the rotating shaft, dipping into the oil sump as it rotates. The ring rotates because of its contact with the shaft, but at a slower speed. The oil adheres to the ring until it reaches the top of the journal when it flows onto the shaft. Figure 11-22. Splash lubrication system (oil stinger). Reciprocating Compressors 315 lo a forced-feed lubrication system, a pump circulates lubricating oil through a cooler and filter to a distribution system that directs the oil to all the bearings and crosshead shoes. Figure 11 -23 is a schematic of a typical system. The details of any one system will vary greatly. Major components and considerations of a forced feed lubrication system are as follows: * Main oil pump - Driven from crankshaft. - Should be sized to deliver 110% of the maximum anticipated flow rate. * Auxiliary pump - Backup for the main oil pump. - Electric motor driven. Figure 11-23. Forced-feed lubrication system. 316 Design of GAS-HANDLING Systems and Facilities - Should start automatically when supply pressure falls below a cer- tain level. • Pre-lube pump - Manual or automatic. - Prevents running bearings dry at start. • Oil cooler - Keeps oil temperature below 165°F. - Can use shell-and-tube exchanger with jacket cooling water or air- cooled exchanger. - Sized for 110% of the maximum anticipated duty. • Oil filter - Dual, full flow, with isolation valves arranged so switching can occur without causing a low-pressure shutdown. - Size should be determined by vendor; in lieu of other information use API 618 requirements. »Overhead day tank - Sized to handle one month of oil consumption. - Should be equipped with a level indicator, • Piping - Stainless steel downstream of filters. - No galvanizing. - No socket welding or other pockets that can accumulate dirt down- stream of filter. - Carbon steel lines should be pickled, passivated, and coated with rust inhibitor. - Lube oil system from pump discharge to the distribution system should be flushed with lube oil at 160°F-180°R Oil should flow across a 200 mesh screen and flushing should cease when no more dirt or grit is found on the screen. Packing/cylinder lubrication can be provided from a forced feed com- pressor lube oil system. For very cold installations, immersion heaters and special lube oils must be considered. If the lube oil temperature gets too cold, the oil becomes too viscous and does not flow and lubricate properly. Cylinder/Packing Lubrication System The flow required to lubricate the packing and cylinders is quite small, and the pressure necessary to inject the lubricant at these locations is quite high. Therefore, small plunger pump (force-feed lubricators) sys- Reciprocating Compressors 317 terns are used. The force-feed lubricators are usually driven by the com- pressor crankshaft, The two basic types of cylinder lubrication systems are the pump-to- point system and the divider-block system. The pump-to-point system provides each lubrication point with its own lubricator pump. Thus, if the compressor cylinders and packing require six lubrication points, the lubri- cator box would be supplied with six cam driven pumps. The divider- block system uses one or more lubricator pumps to supply a divider block, which then distributes the flow to each of the lubrication points. The two systems are sometimes combined such that each stage of compression is provided with its own pump and a divider block to distribute the flow between the cylinders and packing of that particular stage. Oil is supplied to this system from the frame lube oil system or from an overhead tank. This oil comes in contact with and thus contaminates the gas being compressed. Gas/oil compatibility should be checked. PIPE SIZING CONSIDERATIONS Because of the reciprocating action of the piston, care must be exer- cised to size the piping to minimize acoustical pulsations and mechanical vibrations. As a rule of thumb, suction and discharge lines should be sized for a maximum actual velocity of 30 ft/sec (1,800 ft/min) to 42 ft/sec (2,500 ft/min). Volume 1 contains the necessary formulas for deter- mining pressure drop and velocity in gas piping. Analog or digital simulators can be used to establish the pulsation per- formance of any compressor piping system in detail. API 618 Section 3.9.2 provides guidelines for piping pulsation and vibration control based on compressor discharge pressure and horsepower. In practice, many operators do not "analog" compressors of 1,000 horsepower or lower, but rather rely on extrapolations from proven designs. For larger horsepower sizes or where unusual conditions (e.g., unloading and loading cylinders) exist, an analog is recommended. For smaller, high-speed compressors the piping sizing rules of thumb discussed above, in conjunction with pulsation bottles sized from Figure 11 -24, should be sufficient for individual field compressors. These rules of thumb can also be used for preliminary sizing of piping and bottles in preparation for an analog study. To minimize pipe vibrations it is necessary to design pipe runs so that the "acoustic length" of the pipe run does not create a standing wave that 318 Design of GAS-HANDLING Systems and Facilities Figure 11 -24. Pulsation bottle sizing chart (approximation). (Reprinted wilh permission from GPSA Engineering Data Book, 10th Ed.) amplifies the pressure pulsations in the system. The acoustic length is the total overall length from end point to end point including all elbows, bends, and straight pipe runs. Typical pipe runs with respect to acoustic length are considered to be: * Pipe length from suction pipeline to suction scrubber * Pipe length from scrubber to suction pulsation dampeners » Pipe length from discharge pulsation dampeners to cooler « Pipe length from cooler to scrubber * Pipe length from discharge scrubber to pipeline The end of a pipe ran can be classified as either "open" or "closed." Typically, closed ends are where the pipe size is dramatically reduced, as at orifice plates and at short length flow nozzles. A typical open end is where the pipe size is dramatically increased. Where the pipe run contains similar ends (closed-closed or open- open), prohibited pipe lengths are: 0.5X,X 1.5A,,2X where X = acoustic wave length, ft Where the pipe ran contains dissimilar ends (closed-open or open-closed), prohibited pipe lengths are: 0.25X 0.75X, 1.25X, 1.75X The wave length may be calculated from: Reciprocating Compressors 319 where A, = acoustic wavelength, ft k = ratio of specific heats, dimensionless T = gas temperature, °R MW = molecular weight of gas RC = compressor speed, rpm Mechanical vibration of pipe is handled in the same manner as for recip- rocating pumps (Volume 1, Chapter 12). Normally, if the pipe support spacing is kept short, the pipe is securely tied down, the support spans are not uniform in length, and fluid pulsations have been adequately damp- ened, mechanical pipe vibrations will not be a problem. It is good practice to ensure that the natural frequency of all pipe spans is higher than the cal- culated pulsation frequency. The pulsation frequency is given by: where f p = cylinder pulsation frequency, cps n = 1 for single-acting cylinders and n = 2 for double-acting R c = speed of compressor, rpm Refer to Volume 1, Chapters 8 and 9 for the calculations of natural fre- quency of pipe. Foundation Design Considerations Satisafactory compressor installations many times depend on how well the foundation or support structure was designed. An inadequate founda- tion design can result in equipment damage due to excessive vibration. The money saved by cutting corners on foundation design effort may be spent many times in costs associated with high maintenance and lost production. Due to the basic design of the compressor, its rotating and reciprocating masses produce inertia forces and moments tha cannot be completely elim- inated and must be absorbed by the foundation. The manufacturer has the ability to rninimize the magnitude of these forces and moments by adding counterweights to the crossheads but cannot totally eliminate them. 320 Design of GAS-HANDLING Systems and Facilities In addition to the unbalanced forces and moments, the foundation must absorb the moments produced by the gas torque. This is the torque created by the gas pressure forces as the compressor goes through a revo- lution. The compressor manufacturer must provide the magnitude of the resulting forces and moments and the gas torques, Typically foundation design engineers have only used the compressor unbalanced forces and moments in their design calculations. Recent experience has found that the moments created by the gas torque can have a significant impact on foundation design. Detailed information and good design practices for compressor support structures and foundations may be found in Design of Structures and Foundations for Vibrating Machines by Suresh Arya, Michael O'Neill, and George Pincus. For complex offshore structures or where foundations may be critical, finite-element analysis computer programs with dynamic simulation capability can be used to evaluate foundation natural frequency and the forced vibration response. Industry Standard Specifications As previously discussed in this chapter and in Chapter 10, reciprocat- ing compressors are generally classified as either low-speed (integral) compressors or high-speed (separable) compressors. API has provided a standard and specification for each type of compressor to help the user and the facility engineer provide reliable compressor installations. API Standard 618 "Reciprocating Compressors for Petroleum, Chemi- cal, and Gas Industry Services" covers moderate- to low-speed compres- sors in critical services. Integral compressors and low-speed, long stroke balanced-opposed compressors with speeds from 200 to 600 rpm gener- ally fall into this type of construction. The use of this standard with high- speed packaged separable compressors generally results in pages of exceptions by the compressor packager, API Specification IIP "Specification for Packaged Reciprocating Compressors for Oil and Gas Production Services" covers packaged high-speed separable compressors with speeds from 600 to 1,200 rpm. The majority of reciprocating compressors sold in today's market fall into this category. The user and facilitiy engineer must determine the critical nature of each installation and determine the type of construction desired. He or she must consider such things as intended service, compressor location, the conse- quences of downtime, and frequency of up-set or abnormal conditions. [...]... Steel Plates and Sheets Low Alloy Steel Plates SA-516 SA -28 5 SA-36 SA-387 SA -20 3 High Alloy Steel Plates SA - 24 0 Grade 55 Grade 60 Grade 65 Grade 70 Grade A Grade B Grade C Grade 2, cl 1 Grade 12, cl.l Grade 11, cl 1 Grade 22 , cl.l Grade21,cl.l Grade 5, cl.l Grade 2, cl .2 Grade 1 .2, cl .2 Grade 11, cl .2 Grade 22 , cl .2 Grade 21 , cl .2 Grade 5, cl .2 Grade A Grade B Grade D Grade E Grade 3 04 Grade 304L Grade... thickness 2: 1 ellipsoidal heads Wall thickness—hemispherical heads ASME Section Vili Div 1 Div, 2 -20 °F ~20 °F 650°F 100°F 1 3,800 15,000 16,300 17,500 11,300 12, 500 13,800 12, 700 13,800 13,800 15,000 15,000 15,000 13,900 1 7,500 16,300 18,800 17,700 17,700 17 ,40 0 16,300 17,500 16,300 17,500 11 ,20 0 — 12, 300 10 ,20 0 18,300 20 ,000 21 700 23 ,300 15,000 16,700 18,100 1 6,900 18,300 18,300 20 ,000 20 ,000 20 ,000 20 ,000... should have access to a copy of the ASME Code and should become *Reviewed for the 1999 edition by K S Chiou of Paragon Engineering Services, Inc 327 328 Design of GAS-HANDLING Systems and Facilities familiar with its general contents In particular, Section VIII of the code, "Pressure Vessels," is particularly important Countries that do not use the ASME Code have similar documents and requirements The procedures... 150 300 40 0 600 900 1500 25 00 28 5 740 990 148 0 22 20 3705 6170 25 0 675 900 1350 20 25 3375 5 625 Mechanical Design of Pressure Vessels 331 process conditions in their bids and ask the vessel manufacturers to state the maximum MAWP for which the vessel could be tested and approved, Maximum Allowable Stress Values The maximum allowable stress values to be used in the calculation of the vessel's wall thickness... Between Operating and MAWP Less than 50 psig 10 psi 5 1 psig to 25 0 psig 25 psi 25 1 psig to 500 psig 10% of maximum operating pressure 501 psig to 1000 psig 50 psi 1001 psig and higher 5% of maximum operating pressure Vessels with high-pressure safety sensors have an additional 5% or 5 psi, whichever is greater, added to the minimum differential 330 Design of GAS-HANDLING Systems and Facilities accordance... a drawing of the specific cylinder and liner is available However, it should not vary greatly The PD required is: Reciprocating Compressors 325 4 Calculate the rod load The calculated rod load for both the compression and tension modes are within the 25 ,000-Ib maximum rod load limit, 5 Calculate the required horsepower needed for the given conditions: 326 Design of GAS-HANDLING Systems and Facilities. .. 17,500 16,300 17,500 11 ,20 0 — 12, 300 10 ,20 0 18,300 20 ,000 21 700 23 ,300 15,000 16,700 18,100 1 6,900 18,300 18,300 20 ,000 20 ,000 20 ,000 20 ,000 23 ,300 21 ,700 25 ,000 25 ,000 25 ,000 25 .000 21 .700 23 ,300 21 ,700 23 ,300 20 ,000 16,700 20 ,000 16,700 Mechanical Design of Pressure Vessels \X7a1l fht/»l^~ri«aoe 333 r»r»nAC where S = maximum allowable stress value, psi t = thickness, excluding corrosion allowance,... allowance of 0. 125 in for non-corrosive service and 0 .25 0 in for corrosive service is added to the wall thickness calculated in Equations 12- 1 to 12- 4 INSPECTION PROCEDURES All ASME Code vessels are inspected by an approved Code inspector The manufacturer will supply Code papers signed by the inspector The 3 34 Design of GAS-HANDLING Systems and Facilities Figure 12- 1 Pressure vessel shapes nameplate on the... limit of 25 ,000 Ib, and rod diameter of 1.75 in Assume k = 1 .26 , Zs = 0.88, and Zd = 0.85  322 Design of GAS-HANDLING Systems and Facilities Figure 11 -25 Typical compressor cylinder inert buffer gas arrangement (Courtesy ofDresser-Rana Company.) Compute discharge temperature, volumetric efficiency, required clearance, rod load, and required horsepower for the given conditions Also calculate the lowest... fabrication of pressure vessels are given in Table 12- 3 For stress values at higher temperatures and for other materials, the latest edition of the ASME Code should be referenced Determining Wall Thickness The following formulas are used in the ASME Code Section VIII, Division 1 for determining wall thickness: Wall thickness—cylindrical shells 3 32 Design of GAS-HANDLING Systems and Facilities Table 12- 3 Maximum . Material Group 1.1 Class 150 300 40 0 600 900 1500 25 00 ~20 °FtolOO° 28 5 740 990 148 0 22 20 3705 6170 MAWP, psig F 100°Fto200°F 25 0 675 900 1350 20 25 3375 5 625 . temperature 322 Design of GAS-HANDLING Systems and Facilities Reciprocating Compressors 323 2. Calculate the volumetric efficiency 3. Calculate the required clearance Convert to standard. pressure. 3 24 Design of GAS-HANDLING Systems and Facilities (a) Calculate required rpm to give desired throughput: (b) Calculate the clearance that would be needed to reduce the throughput