Section 13 — Compressors and Expanders - Industrial Air Power

SECTION 13Compressors and ExpandersCompressorsDepending on application, compressors are manufacturedas positive-displacement, dynamic, or thermal type (Fig. 13-2).Positive displacement types fall in two basic categories: reciprocatingand rotary.The reciprocating compressor consists of one or more cylinderseach with a piston or plunger that moves back and forth,displacing a positive volume with each stroke.The diaphragm compressor uses a hydraulically pulsed flexiblediaphragm to displace the gas.FIG. 13-1NomenclatureAVR = actual volumetric rate m 3 /h (i.e. at process conditions)Q = inlet capacity (IVR) m 3 /hA p = cross sectional area of piston, (mm 2 )R = universal gas constant = 8.314A r = cross sectional area of piston rod, (mm 2 )kg • mBp = brake or shaft power kW= 213.6kmole • KC = cylinder clearance as a percent of cylinder volumekJC p = specific heat at constant pressure, kJ/(kg • °K)= 8.314kmole • KC v = specific heat at constant volume, kJ/(kg • °K)D = cylinder inside diameter, mmd = piston rod diameter, mmEP = extracted power of expander kWF = an allowance for interstage pressure drop, Eq 13-4G hp= gas power, actual compression power, kW,T = absolute temperature, Kexcluding mechanical losses, WT c = critical temperature, KH = head, N • M/kgT R = reduced temperature, T/T ch = enthalpy, kJ/kgt = temperature, °CIVR = inlet volumetric rate m 3 /h, usually at suctionconditionsk = isentropic exponent, C p /C vMC p = molar specific heat at constant pressure,kJ/(kmole • °K)MC v = molar specific heat at constant volume,kJ/(kmole • °K)M = molecular massN r = speed, rpmN m = molar flow, moles/hn = polytropic exponent or number of molesP = pressure, kPa (abs)P c = critical pressure, kPa (abs)PD = piston displacement, m 3 /hP L = pressure base used in the contract or regulation,kPa (abs)pP c = pseudo critical pressure, kPa (abs)P R = reduced pressure, P/P cpT c = pseudo critical temperature, KP V = partial pressure of contained moisture, kPa (abs)p = pressure, kg/m 2kPa (abs) • m3kg mole • °Kr = compression ratio, P 2 /P 1s = entropy, kJ/kg • K) or number of wheelsSVR = standard volumetric rate m 3 /h measured at101.35 kPa and 15°Cstroke = length of piston movement, mmV = specific volume, m 3 /kgVE = volumetric efficiency, percentW = work, N • Mw = weight flow, kg/hX = temperature rise factory = mole fractionZ = compressibility factorZ avg = average compressibility factor = (Z s + Z d )/2η = efficiency, expressed as a decimalSubscriptsavg = averaged = dischargeis = isentropic processL = standard conditions used for calculation orcontractp = polytropic processS = standard conditions, usually 101.35 kPa (abs), 15°Cs = suctiont = total or overall1 = inlet conditions2 = outlet conditions13-1

FIG. 13-2Types of CompressorsRotary compressors cover lobe-type, screw-type, vane-type,and liquid ring type, each having a casing with one or morerotating elements that either mesh with each other such aslobes or screws, or that displace a fixed volume with each rotation.The dynamic types include radial-flow (centrifugal), axialflow,and mixed flow machines. They are rotary continuousflowcompressors in which the rotating element (impeller orbladed rotor) accelerates the gas as it passes through the element,converting the velocity head into static pressure, partiallyin the rotating element and partially in stationarydiffusers or blades.Ejectors are "thermal" compressors that use a high velocitygas or steam jet to entrain the inflowing gas, then convert thevelocity of the mixture to pressure in a diffuser.Fig. 13-3 covers the normal range of operation for compressorsof the commercially available types. Fig. 13-4 summarizesthe difference between reciprocating and centrifugalcompressors.RECIPROCATING COMPRESSORSReciprocating compressor ratings vary from fractional tomore than 20,000 hp per unit. Pressures range from low vacuumat suction to 30,000 psi and higher at discharge for specialprocess compressors.Reciprocating compressors are furnished either single-stageor multi-stage. The number of stages is determined by theoverall compression ratio. The compression ratio per stage(and valve life) is generally limited by the discharge temperatureand usually does not exceed 4, although small-sized units(intermittent duty) are furnished with a compression ratio ashigh as 8.Gas cylinders are generally lubricated, although a non-lubricateddesign is available when warranted; example: nitrogen,oxygen, and instrument air.On multistage machines, intercoolers may be provided betweenstages. These are heat exchangers which remove theheat of compression from the gas and reduce its temperatureto approximately the temperature existing at the compressorintake. Such cooling reduces the actual volume of gas going tothe high-pressure cylinders, reduces the horsepower requiredfor compression, and keeps the temperature within safe operatinglimits.Reciprocating compressors should be supplied with cleangas as they cannot satisfactorily handle liquids and solid particlesthat may be entrained in the gas. Liquids and solid particlestend to destroy cylinder lubrication and cause excessivewear. Liquids are non-compressible and their presence couldrupture the compressor cylinder or cause other major damage.Performance CalculationsThe engineer in the field is frequently required to:1. determine the approximate horsepower required to compressa certain volume of gas from some intake conditionsto a given discharge pressure, and2. estimate the capacity of an existing compressor underspecified suction and discharge conditions.The following text outlines procedures for making these calculationsfrom the standpoint of quick estimates and also presentsmore detailed calculations. For specific information ona given compressor, consult the manufacturer of that unit.For a compression process, the enthalpy change is the bestway of evaluating the work of compression. If a P-H diagramis available (as for propane refrigeration systems), the workof compression would always be evaluated by the enthalpy13-2

FIG. 13-3Compressor Coverage Chartchange of the gas in going from suction to discharge conditions.Years ago the capability of easily generating P-H diagrams fornatural gases did not exist. The result was that many ways ofestimating the enthalpy change were developed. They wereused as a crutch and not because they were the best way toevaluate compression horsepower requirements.FIG. 13-4Comparison of Reciprocating andCentrifugal CompressorsThe advantages of a centrifugal compressor over a reciprocatingmachine are:a. Lower installed first cost where pressure and volumeconditions are favorable,b. Lower maintenance expense,c. Greater continuity of service and dependability,d. Less operating attention,e. Greater volume capacity per unit of plot area,f. Adaptability to high-speed low-maintenance-cost drivers.The advantages of a reciprocating compressor over a centrifugalmachine are:a. Greater flexibility in capacity and pressure range,b. Higher compressor efficiency and lower power cost,c. Capability of delivering higher pressures,d. Capability of handling smaller volumes,e. Less sensitive to changes in gas composition and density.Today the engineer does have available, in many cases, thecapability to generate that part of the P-H diagram requiredfor compression purposes. This is done using equations of stateon a computer. This still would be the best way to evaluate thecompression horsepower. The other equations are used only ifaccess to a good equation of state is not available.Section 13 continues to treat reciprocating and centrifugalmachines as being different so far as estimation of horsepowerrequirements is concerned. This treatment reflects industrypractice. The only difference in the horsepower evaluation isthe efficiency of the machine. Otherwise the basic thermodynamicequations are the same for all compression.The reciprocating compressor horsepower calculations presentedare based on charts. However, they may equally wellbe calculated using the equations in the centrifugal compressorsection, particularly Eqs. 13-25 through 13-43. This alsoincludes the mechanical losses in Eqs. 13-37 and 13-38.There are two ways in which the thermodynamic calculationsfor compression can be carried out — by assuming:1.isentropic reversible path — a process during which thereis no heat added to or removed from the system and theentropy remains constant, pv k = constant2.polytropic reversible path — a process in which changes ingas characteristics during compression are considered, pv n= constant13-3

Fig. 13-5 shows a plot of pressure vs. volume for each valueof exponent n. The work, W, performed in proceeding from p 1to p 2 along any polytropic curve (Fig. 13-5) isW = ∫21p 2V • dp = ∫ V • dp Eq 13-1p 1The amount of work required is dependent upon thepolytropic curve involved and increases with increasing valuesof n. The path requiring the least amount of input work is n =1, which is equivalent to isothermal compression, a processduring which there is no change in temperature. For isentropiccompression, n = k = ratio of specific heat at constant pressureto that at constant volume.It is usually impractical to build sufficient heat-transferequipment into the design of most compressors to carry awaythe bulk of the heat of compression. Most machines tend tooperate along a polytropic path which approaches the isentropic.Most compressor calculations are therefore based on anefficiency applied to account for true behavior.A compression process following the middle curve in Fig.13-5 has been widely referred to in industry as "adiabatic".However, all compression processes of practical importanceare adiabatic. The term adiabatic does not adequately describethis process, since it only implies no heat transfer. The idealprocess also follows a path of constant entropy and should becalled "isentropic", as will be done subsequently in this chapter.Eq 13-3 which applies to all ideal gases can be used to calculatek.MC p − MC v = R = 1.986 Btu/(lb mol • °F) Eq 13-2By rearrangement and substitution we obtain:k = C pC v= MC pMC v=FIG. 13-5Compression CurvesMC pMC p − 1.986Eq 13-3To calculate k for a gas we need only know the constantpressure molar heat capacity (MC p ) for the gas. Fig. 13-6 givesvalues of molecular weight and ideal-gas state heat capacity(i.e. at 1 atm) for various gases. The heat capacity varies considerablywith temperature. Since the temperature of the gasincreases as it passes from suction to discharge in the compressor,k is normally determined at the average of suctionand discharge temperatures.For a multi-component gas, the mole weighted averagevalue of molar heat capacity must be determined at averagecylinder temperature. A sample calculation is shown in Fig.13-7.The calculation of pP c and pT c in Fig. 13-7 permits calculationof the reduced pressure P R = P/pP c mix and reduced temperatureT R = T/pT c mix. The compressibility Z at T and P canthen be determined using the charts in Section 23.If only the molecular weight of the gas is known and not itscomposition, an approximate value for k can be determinedfrom the curves in Fig. 13-8.Estimating Compressor HorsepowerEq 13-4 is useful for obtaining a quick and reasonable estimatefor compressor horsepower. It was developed for largeslow-speed (300 to 450 rpm) compressors handling gases witha specific gravity of 0.65 and having stage compression ratiosabove 2.5.CAUTION: Compressor manufacturers generally rate theirmachines based on a standard condition of 14.4 psia ratherthan the more common gas industry value of 14.7 psia.Due to higher valve losses, the horsepower requirement forhigh-speed compressors (1000 rpm range, and some up to1800 rpm) can be as much as 20% higher, although this is avery arbitrary value. Some compressor designs do not merit ahigher horsepower allowance and the manufacturers shouldbe consulted for specific applications.Brakehorsepower = (22)⎛ ratio ⎞⎜⎝ stage⎟ (# of stages)(MMcfd)(F)⎠Eq 13-4Where:MMcfd = Compressor capacity referred to 14.4 psia and intaketemperatureF = 1.0 for single-stage compression1.08 for two-stage compression1.10 for three-stage compressionEq 13-4 will also provide a rough estimate of horsepower forlower compression ratios and/or gases with a higher specificgravity, but it will tend to be on the high side. To allow for thistendency use a multiplication factor of 20 instead of 22 forgases with a specific gravity in the 0.8 to 1.0 range; likewise,use a factor in the range of 16 to 18 for compression ratiosbetween 1.5 and 2.0.Curves are available which permit easy estimation of approximatecompression-horsepower requirements. Fig. 13-9 istypical of these curves.Example 13-1 — Compress 2 MMcfd of gas at 14.4 psia andintake temperature through a compression ratio of 9 in a2-stage compressor. What will be the horsepower?Solution StepsRatio per stage = √⎺⎺9 = 3From Eq 13-4 we find the brake horsepower to be:(22) (3) (2) (2) (1.08) = 285 BhpFrom Fig. 13-9, using a k of 1.15, we find the horsepowerrequirement to be 136 Bhp/MMcfd or 272 Bhp. For a k of 1.4,the horsepower requirement would be 147 Bhp/MMcfd or 294total horsepower.13-4

The two procedures give reasonable agreement, particularlyconsidering the simplifying assumptions necessary in reducingcompressor horsepower calculations to such a simple procedure.Detailed CalculationsThere are many variables which enter into the precise calculationof compressor performance. Generalized data asFIG. 13-6Molar Heat Capacity MC p (Ideal-Gas State), Btu/(lb mol • °R)*Data source: Selected Values of Properties of Hydrocarbons, API Research Project 44; MW updated to agree with Fig. 23-2GasChemicalformulaMol wtTemperature0°F 50°F 60°F 100°F 150°F 200°F 250°F 300°FMethane CH 4 16.043 8.23 8.42 8.46 8.65 8.95 9.28 9.64 10.01Ethyne (Acetylene) C 2H 2 26.038 9.68 10.22 10.33 10.71 11.15 11.55 11.90 12.22Ethene (Ethylene) C 2H 4 28.054 9.33 10.02 10.16 10.72 11.41 12.09 12.76 13.41Ethane C 2H 6 30.070 11.44 12.17 12.32 12.95 13.78 14.63 15.49 16.34Propene (Propylene) C 3H 6 42.081 13.63 14.69 14.90 15.75 16.80 17.85 18.88 19.89Propane C 3H 8 44.097 15.65 16.88 17.13 18.17 19.52 20.89 22.25 23.561-Butene (Butlyene) C 4H 8 56.108 17.96 19.59 19.91 21.18 22.74 24.26 25.73 27.16cis-2-Butene C 4H 8 56.108 16.54 18.04 18.34 19.54 21.04 22.53 24.01 25.47trans-2-Butene C 4H 8 56.108 18.84 20.23 20.50 21.61 23.00 24.37 25.73 27.07iso-Butane C 4H 10 58.123 20.40 22.15 22.51 23.95 25.77 27.59 29.39 31.11n-Butane C 4H 10 58.123 20.80 22.38 22.72 24.08 25.81 27.55 29.23 30.90iso-Pentane C 5H 12 72.150 24.94 27.17 27.61 29.42 31.66 33.87 36.03 38.14n-Pentane C 5H 12 72.150 25.64 27.61 28.02 29.71 31.86 33.99 36.08 38.13Benzene C 6H 6 78.114 16.41 18.41 18.78 20.46 22.45 24.46 26.34 28.15n-Hexane C 6H 14 86.177 30.17 32.78 33.30 35.37 37.93 40.45 42.94 45.36n-Heptane C 7H 16 100.204 34.96 38.00 38.61 41.01 44.00 46.94 49.81 52.61Ammonia NH 3 17.0305 8.52 8.52 8.52 8.52 8.52 8.53 8.53 8.53Air 28.9625 6.94 6.95 6.95 6.96 6.97 6.99 7.01 7.03Water H 2O 18.0153 7.98 8.00 8.01 8.03 8.07 8.12 8.17 8.23Oxygen O 2 31.9988 6.97 6.99 7.00 7.03 7.07 7.12 7.17 7.23Nitrogen N 2 28.0134 6.95 6.95 6.95 6.96 6.96 6.97 6.98 7.00Hydrogen H 2 2.0159 6.78 6.86 6.87 6.91 6.94 6.95 6.97 6.98Hydrogen sulfide H 2S 34.08 8.00 8.09 8.11 8.18 8.27 8.36 8.46 8.55Carbon monoxide CO 28.010 6.95 6.96 6.96 6.96 6.97 6.99 7.01 7.03Carbon dioxide CO 2 44.010 8.38 8.70 8.76 9.00 9.29 9.56 9.81 10.05* Exceptions: Air - Keenan and Keyes, Thermodynamic Properties of Air, Wiley, 3rd Printing 1947. Ammonia - Edw. R. Grabl, Thermodynamic Properties ofAmmonia at High Temperatures and Pressures, Petr. Processing, April 1953. Hydrogen Sulfide - J. R. West, Chem. Eng. Progress, 44, 287, 1948.FIG. 13-7Calculation of kExample gas mixtureComponentnameMolfractionyDetermination of mixturemol weightIndividualComponentMol weightMWy • MWDetermination of MC p,Molar heat capacityIndividualComponentMC p @150°F*y • MC p@ 150°FDetermination of pseudo critical pressure, pP c,and temperature, pT cComponentcriticalpressureP c psiay • P cComponentcriticaltemperatureT c°Rmethane 0.9216 16.04 14.782 8.95 8.248 666 615.6 343 316.1ethane 0.0488 30.07 1.467 13.78 0.672 707 34.6 550 26.8propane 0.0185 44.10 0.816 19.52 0.361 616 11.4 666 12.3i-butane 0.0039 58.12 0.227 25.77 0.101 528 2.1 734 2.9n-butane 0.0055 58.12 0.320 25.81 0.142 551 3.0 765 4.2i-pentane 0.0017 72.15 0.123 31.66 0.054 490 0.8 829 1.4Total 1.0000 MW = 17.735 MC p = 9.578 pP c = 667.5 pT c = 363.7MC v = MC p − 1.986 = 7.592 k = MC p/MC v = 9.578/7.592 = 1.26*For values of MC p other than @ 150°F, refer to Fig. 13-6y • T c13-5

FIG. 13-8Approximate Heat-Capacity Ratios of Hydrocarbon GasesFrom molar flow (Nm, mols/min)Q = ⎛ 379.5 • 14.7⎞⎛N m T 1 Z 1 ⎞⎜⎝ 520⎟ ⎜⎠ ⎝P 1 Z ⎟⎠ Eq 13-9LFrom these equations, inlet volume to any stage may be calculatedby using the inlet pressure P 1 and temperature T 1 .Moisture should be handled just as any other component inthe gas.In a reciprocating compressor, effective capacity may be calculatedas the piston displacement (generally in cu ft/min)multiplied by the volumetric efficiency.The piston displacement is equal to the net piston area multipliedby the length of piston sweep in a given period of time.This displacement may be expressed:For a single-acting piston compressing on the outer end only,PD = (stroke) (N) (D2 ) πEq 13-10(4) (1728)= 4.55 (10 −4 ) (stroke) (N) (D 2 )For a single-acting piston compressing on the crank endonly,PD = (stroke) (N) (D2 − d 2 ) πEq 13-11(4) (1728)given in this section are based upon the averaging of manycriteria. The results obtained from these calculations, therefore,must be considered as close approximations to true compressorperformance.CapacityMost gases encountered in industrial compression do notexactly follow the ideal gas equation of state but differ in varyingdegrees. The degree in which any gas varies from the idealis expressed by a compressibility factor, Z, which modifies theideal gas equation:PV = nRT Eq 13-5to PV = nZRT Eq 13-6Compressibility factors can be determined from charts inSection 23 using the pP R and pT R of the gas mixture. For purecomponents such as propane, compressibility factors can bedetermined from the P-H diagrams, although the user wouldbe better advised to determine the compression horsepowerusing the P-H diagram (see Section 24).For the purpose of performance calculations, compressor capacityis expressed as the actual volumetric quantity of gas atthe inlet to each stage of compression on a per minute basis(ICFM).From SCFMQ = SCFM ⎛ 14.7⎞⎛T 1 Z 1 ⎞⎜⎝ 520⎟ ⎜⎠ ⎝P 1 Z ⎟⎠ Eq 13-7LFrom weight flow (w, lb/min)Q = 10.73MW⎛⎜⎝w T 1 Z 1P 1 Z L⎞⎟⎠ Eq 13-8= 4.55 (10 −4 ) (stroke) (N) (D 2 − d 2 )For a double-acting piston (other than tail rod type),PD = (stroke) (N) (2 D2 − d 2 ) πEq 13-12(4) (1728)= 4.55 (10 −4 ) (stroke) (N) (2 D 2 − d 2 )Volumetric EfficiencyIn a reciprocating compressor, the piston does not travelcompletely to the end of the cylinder at the end of the dischargestroke. Some clearance volume is necessary and it includes thespace between the end of the piston and the cylinder headwhen the piston is at the end of its stroke. It also includes thevolume in the valve ports, the volume in the suction valveguards, and the volume around the discharge valve seats.Clearance volume is usually expressed as a percent of pistondisplacement and referred to as percent clearance, or cylinderclearance, C.clearance volume, cu in.C =(100) Eq 13-13piston displacement, cu in.For double acting cylinders, the percent clearance is basedon the total clearance volume for both the head end and thecrank end of a cylinder. These two clearance volumes are notthe same due to the presence of the piston rod in the crank endof the cylinder. Sometimes additional clearance volume (external)is intentionally added to reduce cylinder capacity.The term "volumetric efficiency" refers to the actual pumpingcapacity of a cylinder compared to the piston displacement.Without a clearance volume for the gas to expand and delaythe opening of the suction valve(s), the cylinder could deliverits entire piston displacement as gas capacity. The effect of thegas contained in the clearance volume on the pumping capacityof a cylinder can be represented by:VE = 100 − r − C ⎡ ⎢⎣Z sZ d(r 1/k ) − 1 ⎤ ⎥⎦Eq 13-1413-6

FIG. 13-9Approximate Horsepower Required to Compress GasesVolumetric efficiencies as determined by Eq. 13-14 are theoreticalin that they do not account for suction and dischargevalve losses. The suction and discharge valves are actuallyspring-loaded check valves that permit flow in one directiononly. The springs require a small differential pressure to open.For this reason, the pressure within the cylinder at the end ofthe suction stroke is lower than the line suction pressure and,likewise, the pressure at the end of the discharge stroke ishigher than line discharge pressure.One method for accounting for suction and discharge valvelosses is to reduce the volumetric efficiency by an arbitraryamount, typically 4%, thus modifying Eq. 13-14 as follows:VE = 96 − r − C ⎡ ⎢⎣Z sZ d(r 1/k ) − 1 ⎤ ⎥⎦Eq 13-15When a non-lubricated compressor is used, the volumetricefficiency should be corrected by subtracting an additional 5%for slippage of gas. This is a capacity correction only and, as afirst approximation, would not be considered when calculatingcompressor horsepower. The energy of compression is used bythe gas even though the gas slips by the rings and is not dischargedfrom the cylinder.If the compressor is in propane, or similar heavy gas service,an additional 4% should be subtracted from the volumetricefficiency. These deductions for non-lubricated and propaneperformance are both approximate and, if both apply, cumulative.Fig. 13-10 provides the solution to the function r 1/k . Valuesfor compression ratios not shown may be obtained by interpolation.The closest k value column may be safely used withouta second interpolation.Volumetric efficiencies for "high speed" separable compressorsin the past have tended to be slightly lower than estimatedfrom Eq 13-14. Recent information suggests that thismodification is not necessary for all models of high speed compressors.In evaluating efficiency, horsepower, volumetric efficiency,etc., the user should consider past experience with different13-7

FIG. 13-10Values of r 1/kCompressionRatiok, isentropic exponent Cp/Cv1.10 1.14 1.18 1.22 1.26 1.30 1.34 1.38 1.421.2 1.180 1.173 1.167 1.161 1.156 1.151 1.146 1.141 1.1371.4 1.358 1.343 1.330 1.318 1.306 1.295 1.285 1.276 1.2671.6 1.533 1.510 1.489 1.470 1.452 1.436 1.420 1.406 1.3921.8 1.706 1.675 1.646 1.619 1.594 1.572 1.551 1.531 1.5132.0 1.878 1.837 1.799 1.765 1.733 1.704 1.677 1.652 1.6292.2 2.048 1.997 1.951 1.908 1.870 1.834 1.801 1.771 1.7422.4 2.216 2.155 2.100 2.050 2.003 1.961 1.922 1.886 1.8522.6 2.384 2.312 2.247 2.188 2.135 2.086 2.040 1.999 1.9602.8 2.550 2.467 2.393 2.326 2.264 2.208 2.156 2.109 2.0653.0 2.715 2.621 2.537 2.461 2.391 2.328 2.270 2.217 2.1683.2 2.879 2.774 2.680 2.595 2.517 2.447 2.382 2.323 2.2693.4 3.042 2.926 2.821 2.727 2.641 2.563 2.492 2.427 2.3673.6 3.204 3.076 2.961 2.857 2.764 2.679 2.601 2.530 2.4653.8 3.366 3.225 3.100 2.987 2.885 2.792 2.708 2.631 2.5604.0 3.526 3.374 3.238 3.115 3.005 2.905 2.814 2.731 2.6554.2 3.686 3.521 3.374 3.242 3.124 3.016 2.918 2.829 2.7474.4 3.846 3.668 3.510 3.368 3.241 3.126 3.021 2.926 2.8394.6 4.004 3.814 3.645 3.493 3.357 3.235 3.123 3.022 2.9294.8 4.162 3.959 3.779 3.617 3.473 3.342 3.224 3.116 3.0185.0 4.319 4.103 3.912 3.740 3.587 3.449 3.324 3.210 3.1065.2 4.476 4.247 4.044 3.863 3.700 3.554 3.422 3.303 3.1935.4 4.632 4.390 4.175 3.984 3.813 3.659 3.520 3.394 3.2795.6 4.788 4.532 4.306 4.105 3.925 3.763 3.617 3.485 3.3645.8 4.943 4.674 4.436 4.224 4.035 3.866 3.713 3.574 3.4486.0 5.098 4.815 4.565 4.343 4.146 3.968 3.808 3.663 3.5326.2 5.252 4.955 4.694 4.462 4.255 4.069 3.902 3.751 3.6146.4 5.406 5.095 4.822 4.579 4.363 4.170 3.996 3.839 3.6966.6 5.560 5.235 4.949 4.696 4.471 4.270 4.089 3.925 3.7776.8 5.713 5.374 5.076 4.813 4.578 4.369 4.181 4.011 3.8577.0 5.865 5.512 5.202 4.928 4.685 4.468 4.272 4.096 3.937speeds and models. Larger valve area for a given swept volumewill generally lead to higher compression efficiencies.Equivalent CapacityThe net capacity for a compressor, in cubic feet per day @14.4 psia and suction temperature, may be calculated byEq. 13-16a which is shown in dimensioned form:MMcfd =PD ft3 min• 1440 VE%min d• 100 • P lbsin14.4 lbin 2 • Z s2• 10−6MMft3ft 3 • Z 14.4Eq 13-16awhich can be simplified to Eq. 13-16b when Z 14.4 is assumedto equal 1.0.MMcfd = PD • VE • P s • 10 −6Z sEq 13-16bFor example, a compressor with 200 cu ft/min piston displacement,a volumetric efficiency of 80%, a suction pressureof 75 psia, and suction compressibility of 0.9 would have a capacityof 1.33 MMcfd at 14.4 psia and suction temperature. Ifcompressibility is not used as a divisor in calculating cu ft/min,then the statement "not corrected for compressibility" shouldbe added.In many instances the gas sales contract or regulation willspecify some other measurement standard for gas volume. Toconvert volumes calculated using Equation 13-16 (i.e. at14.4 psia and suction temperature) to a P L and T L basis,Eq 13-17 would be used:MMscfd at P L , T L = (MMcfd from Eq 13−16) ⎛ 14.4⎞ ⎛T L ⎞ ⎛Z L ⎞⎜⎝P ⎟⎠ ⎜⎝L T ⎟⎠ ⎜⎝s Z ⎟⎠sDischarge TemperatureEq 13-17The temperature of the gas discharged from the cylinder canbe estimated from Eq 13-18, which is commonly used but notrecommended. (Note: the temperatures are in absolute units,°R or K.) Eqs 13-31 and 13-32 give better results.T d = T s (r (k − 1)/k ) Eq 13-18Fig. 13-11 is a nomograph which can be used to solveEq 13-18. The discharge temperature determined from either13-8

Eq 13-18 or Fig. 13-11 is the theoretical value. While it neglectsheat from friction, irreversibility effects, etc., and maybe somewhat low, the values obtained from this equation willbe reasonable field estimates.Rod LoadingEach compressor frame has definite limitations as to maximumspeed and load-carrying capacity. The load-carryingcapacity of a compressor frame involves two primary considerations:horsepower and rod loading.The horsepower rating of a compressor frame is the measureof the ability of the supporting structure and crankshaft towithstand torque (turning force) and the ability of the bearingsto dissipate frictional heat. Rod loads are established tolimit the static and inertial loads on the crankshaft, connectingrod, frame, piston rod, bolting, and projected bearing surfaces.Good design dictates a reversal of rod loading during eachstroke. Non-reversal of the loading results in failure to allowbearing surfaces to part and permit entrance of sufficient lubricant.The result will be premature bearing wear or failure.Rod loadings may be calculated by the use of Eqs 13-19 and13-20.Load in compression = P d A p − P s (A p − A r )Load in tension = P d (A p − A r ) − P s A p= (P d − P s ) A p + P s A r Eq 13-19= (P d − P s ) A p − P d A r Eq 13-20Using Eqs. 13-19 and 13-20, a plus value for the load in bothcompression and tension indicates a reversal of loads based ongas pressure only. Inertial effects will tend to increase the degreeof reversal.The true rod loads would be those calculated using internalcylinder pressures after allowance for valve losses. Normally,the operator will know only line pressures, and because of this,manufacturers generally rate their compressors based on linepressurecalculations.A further refinement in the rod-loading calculation wouldbe to include inertial forces. While the manufacturer will considerinertial forces when rating compressors, useful data onthis point is seldom available in the field. Except in specialcases, inertial forces are ignored.A tail-rod cylinder would require consideration of rod crosssectionarea on both sides of the piston instead of on only oneside of the piston, as in Eqs 13-19 and 13-20.HorsepowerDetailed compressor horsepower calculations can be madethrough the use of Figs. 13-12 and 13-13. For ease of calculations,these figures provide net horsepower, including mechanicalefficiency and gas losses. Figs. 13-14 and 13-15 areincluded for modifying the horsepower numbers for specialconditions.Proper use of these charts should provide the user with reasonablycorrect horsepower requirements that are comparableto those calculated by the compressor manufacturer. For moredetailed design, the engineer should consult a compressormanufacturer.Volumes to be handled in each stage must be corrected tothe actual temperature at the inlet to that stage. Note thatmoisture content corrections can also be important at lowpressure and/or high temperature.When intercoolers are used, allowance must be made forinterstage pressure drop. Interstage pressures may be estimatedby:1. Obtaining the overall compression ratio, r t .2. Obtaining the calculated ratio per stage, r, by taking thes root of r t , where s is the number of compression stages.3. Multiplying r by the absolute intake pressure of the stagebeing considered.This procedure gives the absolute discharge pressure of thisstage and the theoretical absolute intake pressure to the nextstage. The next stage intake pressure can be corrected for intercoolerpressure drop by reducing the pressure by 3-5 psi.This can be significant in low pressure stages.Horsepower for compression is calculated by using Figs.13-12 and 13-13 and Eq. 13-21.Bhp = (Bhp/MMcfd) ⎛ P L ⎞ ⎛ ⎞⎜ ⎟⎝ 14.4⎜⎠ ⎝T ⎟⎠ (Z avg ) (MMcfd)LEq 13-21Bhp/MMcfd is read from Figs. 13-12 and 13-13 which use apressure base of 14.4 psia.T s13-9

FIG. 13-11Theoretical Discharge TemperaturesSingle-Stage CompressionRead r to k to t s to t d13-10

FIG. 13-12Bhp Per Million CurveMechanical Efficiency-95%Gas Velocity Through Valve-3000 ft/min (API equation)13-11

FIG. 13-13Bhp Per Million CurveMechanical Efficiency-95%Gas Velocity Through Valve-3000 ft/min (API equation)13-12

FIG. 13-14Correction Factor for Low Intake PressureFIG. 13-15Correction Factor for Specific GravityFigs. 13-12 and 13-13 are for standard valved cylinders.Caution should be used in applying conventional cylinders tolow compression-ratio pipeline compressors. For low ratiopipeline compressors a high clearance type cylinder permitsvalve designs with higher efficiency. The compressor manufacturershould be consulted for Bhp curves on this type cylinder.Fig. 13-14 provides a correction for intake pressure. The correctionfactor, as read from the curve, is used as a multiplierin the right hand side of Eq 13-21 to obtain the corrected brakehorsepower.Fig. 13-15 provides a correction factor for gas specific gravity.The correction factor is used as a multiplier in the righthandside of Eq 13-21 to obtain the corrected horsepower.Data presented in Figs. 13-12 and 13-13 are for slow speedintegral compressors rather than the high speed separablecompressors. To adjust the horsepower for the high speed unit,the values obtained from Figs. 13-12 and 13-13 may be increasedby the following percentages:Gas SpecificGravity0.5-0.8 40.9 51.0 61.1 81.5and propanerefrigeration units 10Percent horsepowerincrease forhigh speed unitsBecause of variations by different manufacturers in specifyingvalve velocities for high speed as opposed to slow speedcompressors, a given unit may differ from the horsepower correctionsshown. Experience with compressors from a specificmanufacturer will serve to guide the user and give confidencein utilization of the correction factors shown. For applicationswhich are outside typical ranges discussed here, compressormanufacturers should be consulted.Example 13-2 — Compress 2 MMscfd of gas measured at14.65 psia and 60°F. Intake pressure is 100 psia, and intaketemperature is 100°F. Discharge pressure is 900 psia. The gashas a specific gravity of 0.80 (23 MW). What is the requiredhorsepower?1. Compression ratio is900 psia100 psia = 9This would be a two-stage compressor; therefore, the ratioper stage is √⎺⎺9 or 3.2. 100 psia x 3 = 300 psia (1st stage discharge pressure)300 psia – 5 = 295 psia (suction to 2nd stage)Where the 5 psi represents the pressure drop betweenfirst stage discharge and second stage suction.900 psia295 psia= 3.05 (compression ratio for 2nd Stage)It may be desirable to recalculate the interstage pressureto balance the ratios. For this sample problem, however,the first ratios determined will be used.13-13

3. From Fig. 13-8 a gas with specific gravity of 0.8 at 150°Fwould have an approximate k of 1.21. For most compressionapplications, the 150°F curve will be adequate. Thisshould be checked after determining the average cylindertemperature.4. Discharge temperature for the 1st stage may be obtainedby using Fig. 13-11 or solving Eq 13-18. For a compressionratio of 3, discharge temperature = approximately220°F. Average cylinder temperature = 160°F.5. In the same manner, discharge temperature for the secondstage (with r = 3.05 and assuming interstage coolingto 120°F) equals approximately 244°F. Average cylindertemperature = 182°F.6. From the physical properties section (Section 23), estimatethe compressibility factors at suction and dischargepressure and temperature of each stage.1st stage: Z s = 0.98Z d = 0.97Z avg = 0.9752nd stage: Z s = 0.94Z d = 0.92Z avg = 0.937. From Fig. 13-12, Bhp/MMcfd at 3 compression ratio anda k of 1.21 is 63.5 (1st stage).From Fig. 13-12, Bhp/MMcfd at 3.05 compression ratioand a k of 1.21 is 64.5 (second stage).8. There are no corrections to be applied from Fig. 13-14 or13-15, as all factors read unity.9. Substituting in Eq 13-21:1st stage:Bhp/MMscfd = 63.5 ⎛ ⎜⎝14.6514.4⎞ ⎛560⎞⎟ ⎜⎠ ⎝520⎟ 0.975 = 67.8⎠Bhp for 1st stage = 2 MMscfd × 67.8 = 135.62nd stage:Bhp/MMscfd = 64.5 ⎛ ⎜⎝14.6514.4⎞ ⎛580⎞⎟ ⎜⎠ ⎝520⎟ 0.93 = 68.1⎠Bhp for 2nd stage = (2 MMscfd) (68.1) = 136.2Total Bhp for this application = 135.6 + 136.2 = 271.8.Note that in Example 13-1 the same conditions result in acompression power of 285 Bhp which is close agreement.Limits to compression ratio per stage — The maximumratio of compression permissible in one stage is usuallylimited by the discharge temperature or by rod loading, particularlyin the first stage.When handling gases containing oxygen, which could supportcombustion, there is a possibility of fire and explosionbecause of the oil vapors present.To reduce carbonization of the oil and the danger of fires, asafe operating limit may be considered to be approximately300°F. Where no oxygen is present in the gas stream, temperaturesof 350°F may be considered as the maximum, eventhough mechanical or process requirements usually dictate alower figure.Packing life may be significantly shortened by the dual requirementto seal both high pressure and high temperaturegases. For this reason, at higher discharge pressures, a temperaturecloser to 250°F or 275°F may be the practical limit.In summary, and for most field applications, the use of 300°Fmaximum would be a good average. Recognition of the abovevariables is, however, still useful.Economic considerations are also involved because a highratio of compression will mean a low volumetric efficiency andrequire a larger cylinder to produce the same capacity. For thisreason a high rod loading may result and require a heavierand more expensive frame.Where multi-stage operation is involved, equal ratios ofcompression per stage are used (plus an allowance for pipingand cooler losses if necessary) unless otherwise required byprocess design. For two stages of compression the ratio perstage would approximately equal the square root of the totalcompression ratio; for three stages, the cube root, etc. In practice,especially in high-pressure work, decreasing the compressionratio in the higher stages to reduce excessive rodloading may prove to be advantageous.Cylinder DesignDepending on the size of the machine and the number ofstages, reciprocating compressors are furnished with cylindersfitted with either single- or double-acting pistons, see examplesin Figs. 13-16 through 13-18.In the same units, double-acting pistons are commonly usedin the first stages and often single-acting in the higher stagesof compression.Cylinder materials are normally selected for strength; however,thermal shock, mechanical shock, or corrosion resistancemay also be a determining factor. The table below shows dischargepressure limits generally used in the gas industry forcylinder material selection.Cylinder MaterialDischarge Pressure (psig)Cast Iron up to 1,200Nodular Iron about 1,500Cast Steel 1,200 to 2,500Forged Steel above 2,500API standard 618 recommends 1000 psig as the maximumpressure for both cast iron and nodular iron.Cylinders are designed both as a solid body (no liner) andwith liners. Cylinder liners are inserted into the cylinder bodyto either form or line the pressure wall. There are two types.The wet liner forms the pressure wall as well as the inside wallof the water jacket. The dry type lines the cylinder wall and isnot required to add strength.Standard cylinder liners are cast iron. If cylinders are requiredto have special corrosion or wear resistance, other materialsor special alloys may be needed.Most compressors use oils to lubricate the cylinder with amechanical, force-feed lubricator having one or more feeds toeach cylinder.The non-lubricated compressor has found wide applicationwhere it is desirable or essential to compress air or gas withoutcontaminating it with lubricating oil.13-14

FIG. 13-16Low Pressure Cylinder with Double-Acting PistonFor such cases a number of manufacturers furnish a "nonlubricated"cylinder (Fig. 13-19). The piston on these cylindersis equipped with piston rings of graphitic carbon or plastic aswell as pads or rings of the same material to maintain theproper clearance between the piston and cylinder. Plasticpacking of a type that requires no lubricant is used on thestuffing box. Although oil-wiper rings are used on the pistonrod where it leaves the compressor frame, minute quantitiesof oil might conceivably enter the cylinder on the rod. Whereeven such small amounts of oil are objectionable, an extendedcylinder connecting piece can be furnished. This simplylengthens the piston rod so that no lubricated portion of therod enters the cylinder.A small amount of gas leaking through the packing can beobjectionable. Special distance pieces are furnished betweenthe cylinder and frame, which may be either single-compartmentor double-compartment. These may be furnished gastight and vented back to the suction, or may be filled with asealing gas or fluid and held under a slight pressure, or simplyvented.Compressor valves for non-lubricated service operate in anenvironment that has no lubricant in the gas or in the cylinder.Therefore, the selection of valve materials is important to preventexcessive wear.Piston rod packing universally used in non-lubricated compressorsis of the full-floating mechanical type, consisting of acase containing pairs of either carbon or plastic (TFE) rings ofconventional design.When handling oxygen and other gases such as nitrogen andhelium, it is absolutely necessary that all traces of hydrocarbonsin cylinders be removed. With oxygen, this is required forsafety, with other gases to prevent system contamination.Deoiling schemes are discussed in Refrigeration, Section 14.High-pressure compressors with discharge pressures from5,000 to 30,000 psi usually require special design and a completeknowledge of the characteristics of the gas.As a rule, inlet and discharge gas pipe connections on thecylinder are fitted with flanges of the same rating for the followingreasons:13-15

FIG. 13-17High Pressure Cylinder with Double-Acting Pistonand Tail-RodFIG. 13-19Piston Equipped with Teflon ® Piston and Wear Rings for aSingle-Acting Non-Lubricated CylinderFIG. 13-18Single-Acting Plunger Cylinder Designed for15,000 psig DischargeOutput of compressors must be controlled (regulated) tomatch system demand.In many installations some means of controlling the outputof the compressor is necessary. Often constant flow is requireddespite variations in discharge pressure, and the control devicemust operate to maintain a constant compressor capacity.Compressor capacity, speed, or pressure may be varied in accordancewith the requirements. The nature of the control devicewill depend on the regulating variable — whetherpressure, flow, temperature, or some other variable — and ontype of compressor driver.Unloading for Starting — Practically all reciprocatingcompressors must be unloaded to some degree before startingso that the driver torque available during acceleration is notexceeded. Both manual and automatic compressor startup unloadingis used. Common methods of unloading include: dischargeventing, discharge to suction bypass, and holding openthe inlet valves using valve lifters.Capacity Control — The most common requirement isregulation of capacity. Many capacity controls, or unloadingdevices, as they are usually termed, are actuated by the pressureon the discharge side of the compressor. A falling pressureindicates that gas is being used faster than it is being compressedand that more gas is required. A rising pressure indicatesthat more gas is being compressed than is being usedand that less gas is required.• Practicality and uniformity of casting and machinery,• Hydrostatic test, usually at 150% design pressure,• Suction pulsation bottles are usually designed for thesame pressure as the discharge bottle (often federal,state, or local government regulation).Reciprocating Compressor Control DevicesA common method of controlling the capacity of a compressoris to vary the speed. This method is applicable to steamdrivencompressors and to units driven by internal combustionengines. In these cases the regulator actuates the steam-admissionor fuel-admission valve on the compressor driver tocontrol the speed.Electric motor-driven compressors usually operate at constantspeed, and other methods of controlling the capacity arenecessary. On reciprocating compressors up to about 100 hp,two types of control are usually available. These are automatic-start-and-stopcontrol and constant-speed control.Automatic-start-and-stop control, as its name implies, stopsor starts the compressor by means of a pressure-actuatedswitch as the gas demand varies. It should be used only whenthe demand for gas will be intermittent.13-16

Constant-speed control permits the compressor to operateat full speed continuously, loaded part of the time and fully orpartially unloaded at other times. Two methods of unloadingthe compressor with this type of control are in common use:inlet-valve unloaders, and clearance unloaders. Inlet-valveunloaders (Fig. 13-20) operate to hold the compressor inletvalves open and thereby prevent compression. Clearance unloaders(Fig. 13-21) consist of pockets or small reservoirswhich are opened when unloading is desired. The gas is compressedinto them on the compression stroke and expands intothe cylinder on the return stroke, reducing the intake of additionalgas.Motor-driven reciprocating compressors above 100 hp insize are usually equipped with a step control. This is in realitya variation of constant-speed control in which unloading isaccomplished in a series of steps, varying from full load downto no load.Five-step control (full load, three-quarter load, one-halfload, one-quarter load, and no load) is accomplished by meansof clearance pockets. On some makes of machines inlet-valveand clearance control unloading are used in combination.A common practice in the natural gas industry is to preparea single set of curves for a given machine unless there are sideloads or it is a multi-service machine.FIG. 13-20Inlet Valve UnloaderFig. 13-22 shows indicator cards which demonstrate the unloadingoperation for a double acting cylinder at three capacitypoints. The letters adjacent to the low-pressure diagrams representthe unloading influence of the respective and cumulativeeffect of the various pockets as identified in Fig. 13-21.Full load, one-half, and no load capacity is obtained by holdingcorresponding suction valves open or adding sufficient clearanceto produce a zero volumetric efficiency. No-capacity operationincludes holding all suction valves open.Fig. 13-23 shows an alternative representation of compressorunloading operation with a step-control using fixed volumeclearance pockets. The curve illustrates the relationship betweencompressor capacity and driver capacity for a varyingcompressor suction pressure at a constant discharge pressureand constant speed. The driver can be a gas engine or electricmotor.The purpose of this curve is to determine what steps of unloadingare required to prevent the driver and piston rods fromserious overloading. All lines are plotted for a single stage compressor.The driver capacity line indicates the maximum allowablecapacity for a given horsepower. The cylinder capacity linesrepresent the range of pressures calculated with all possiblecombinations of pockets open and cylinder unloading, as necessary,to cover the capacity of the driver.Starting at the end (line 0-0) with full cylinder capacity, theline is traced until it crosses the driver capacity line at whichpoint it is dropped to the next largest cylinder capacity andfollow until it crosses the driver line, etc. This will produce a"saw tooth" effect, hence the name "saw tooth" curve. Thenumber of "teeth" depends upon the number of combinationsof pockets (opened or closed) required for unloading.The same method is followed for multi-stage units. For eachadditional stage another "saw tooth" curve must be constructed,i.e., for a two stage application, two curves are requiredto attain the final results.FIG. 13-22Indicator Diagram for Three Load Points of OperationFIG. 13-21Pneumatic Valves Controlling Four Fixed Pockets inCompressor for Five-Step Control13-17

FIG. 13-23"Saw Tooth" Curve for Unloading OperationAlthough control devices are often automatically operated,manual operation is satisfactory for many services. Wheremanual operation is provided, it often consists of a valve, orvalves, to open and close clearance pockets. In some cases, amovable cylinder head is provided for variable clearance in thecylinder (Fig. 13-24).Gas Pulsation ControlPulsation is inherent in reciprocating compressors becausesuction and discharge valves are open during only part of thestroke.Pulsation must be damped (controlled) in order to:a. provide smooth flow of gas to and from the compressor,b. prevent overloading or underloading of the compressors,andc. reduce overall vibration.There are several types of pulsation chambers. The simplestone is a volume bottle, or a surge drum, which is a pressurevessel, unbaffled internally and mounted on or very near acylinder inlet or outlet.A manifold joining the inlet and discharge connections ofcylinders operating in parallel can also serve as a volume bottle.Performance of volume bottles is not normally guaranteedwithout an analysis of the piping system from the compressorto the first process vessel.Volume bottles are sized empirically to provide an adequatevolume to absorb most of the pulsation. Several industrymethods were tried in an effort to produce a reasonable ruleof-thumbfor their sizing. Fig. 13-25 may be used for approximatebottle sizing.Example 13-3Indicated suction pressure = 600 psiaIndicated discharge pressure = 1,400 psiaCylinder bore = 6 in.Cylinder stroke = 15 in.Swept volume = π (6 2 /4) (15) = 424 cu in.From Fig. 13-25:At 600 psi inlet pressure, the suction bottle multiplier is approximately7.5. Suction-bottle volume = (7.5) (424) =3,180 cu in.At 1,400 psi discharge pressure, the discharge bottle multiplieris approximately 8.5. Discharge-bottle volume = (8.5)(424) = 3,600 cu in.NOTE: When more than one cylinder is connected to a bottle,the sum of the individual swept volumes is the size requiredfor the common bottle.For more accurate sizing, compressor manufacturers can beconsulted. Organizations which provide designs and/or equipmentfor gas-pulsation control are also available.FIG. 13-24Sectional View of a Cylinder Equipped with a Hand-Operated Valve Lifter and Variable-Volume Clearance13-18

FIG. 13-25Approximate Bottle Sizing ChartHaving determined the necessary volume of the bottle, theproportioning of diameter and length to provide this volumerequires some ingenuity and judgment. It is desirable thatmanifolds be as short and of as large diameter as is consistentwith pressure conditions, space limitations, and appearance.A good general rule is to make the manifold diameter 1-1/2times the inside diameter of the largest cylinder connected toit, but this is not always practicable, particularly where largecylinders are involved.Inside diameter of pipe must be used in figuring manifolds.This is particularly important in high-pressure work and insmall sizes where wall thickness may be a considerable percentageof the cross sectional area. Minimum manifold lengthis determined from cylinder center distances and connectingpipe diameters. Some additions must be made to the minimumthus determined to allow for saddle reinforcements and forwelding of caps.It is customary to close the ends of manifolds with weldingcaps which add both volume and length. Fig. 13-26 gives approximatevolume and length of standard caps.Pulsation Dampeners (Snubbers)A pulsation dampener is an internally-baffled device. Thedesign of the pulsation dampening equipment is based onacoustical analog evaluation which takes into account thespecified operating speed range, conditions of unloading, andvariations in gas composition.Analog evaluation is accomplished with an active analogthat simulates the entire compressor, pulsation dampeners,piping and equipment system and considers dynamic interactionsamong these elements.Pulsation dampeners also should be mounted as close aspossible to the cylinder, and in large volume units, nozzlesshould be located near the center of the chamber to reduceunbalanced forces.Pulsation dampeners are typically guaranteed for a maximumresidual peak-to-peak pulsation pressure of 2% of averageabsolute pressure at the point of connection to the pipingsystem, and pressure drop through the equipment of not morethan 1% of the absolute pressure. This applies at design conditionand not necessarily for other operating pressures andflows. A detailed discussion of recommended design approachesfor pulsation suppression devices is presented in APIFIG. 13-26Welding CapsStandard weight Extra strong Double Extra strongPipe sizeVolume, cu in. Length, in. Volume, cu in. Length, in. Volume, cu in. Length, in.4" 24.2 2 1 ⁄ 2 20.0 2 1 ⁄ 2 15 36" 77.3 3 1 ⁄ 2 65.7 3 1 ⁄ 2 48 48" 148.5 4 11 ⁄ 16 122.3 4 11 ⁄ 16 120 510" 295.6 5 3 ⁄ 4 264.4 8 3 ⁄ 412" 517.0 6 7 ⁄ 8 475.0 6 7 ⁄ 814" 684.6 7 13 ⁄ 16 640.0 7 13 ⁄ 1616" 967.6 9 911.0 918" 1432.6 10 1 ⁄ 16 1363.0 10 1 ⁄ 1620" 2026.4 11 1 ⁄ 4 1938.0 11 1 ⁄ 424" 3451.0 13 7 ⁄ 16 3313.0 13 7 ⁄ 1613-19

FIG. 13-27Probable Causes of Reciprocating Compressor TroubleTroubleProbable Cause(s)TroubleProbable Cause(s)1. Power supply failure.1. Lubrication failure.Compressor Willnot Start2. Switchgear or starting panel.3. Low oil pressure shut downswitch.Packing Over-Heating2. Improper lube oil and/orinsufficient lube rate.3. Insufficient cooling.Motor Will NotSynchronize4. Control panel.1. Low voltage.2. Excessive starting torque.3. Incorrect power factor.4. Excitation voltage failure.1. Oil pump failure.2. Oil foaming from counterweightsstriking oil surface.3. Cold oil.4. Dirty oil filter.Excessive CarbonOn Valves1. Excessive lube oil.2. Improper lube oil (too light, highcarbon residue).3. Oil carryover from inlet system orprevious stage.4. Broken or leaking valves causinghigh temperature.5. Excessive temperature due tohigh pressure ratio acrosscylinders.5. Interior frame oil leaks.1. Faulty relief valve.Low Oil Pressure6. Excessive leakage at bearingshim tabs and/or bearings.7. Improper low oil pressure switchsetting.8. Low gear oil pump by-pass/reliefvalve setting.9. Defective pressure gauge.10. Plugged oil sump strainer.Relief ValvePopping2. Leaking suction valves or ringson next higher stage.3. Obstruction (foreign material,rags), blind or valve closed indischarge line.1. Excessive ratio on cylinder due toleaking inlet valves or rings onnext higher stage.11. Defective oil relief valve.1. Loose piston.2. Piston hitting outer head orframe end of cylinder.High DischargeTemperature2. Fouled intercooler/piping.3. Leaking discharge valves orpiston rings.4. High inlet temperature.3. Loose crosshead lock nut.5. Fouled water jackets on cylinder.Noise In Cylinder4. Broken or leaking valve(s).5. Worn or broken piston rings orexpanders.6. Valve improperly seated/damagedseat gasket.7. Free air unloader plungerchattering.1. Worn packing rings.Frame Knocks6. Improper lube oil and/or lube rate.1. Loose crosshead pin, pin caps orcrosshead shoes.2. Loose/worn main, crankpin orcrosshead bearings.3. Low oil pressure.4. Cold oil.2. Improper lube oil and/orinsufficient lube rate (blue rings).3. Dirt in packing.5. Incorrect oil.6. Knock is actually from cylinderend.Excessive PackingLeakage4. Excessive rate of pressureincrease.5. Packing rings assembledincorrectly.Crankshaft OilSeal Leaks1. Faulty seal installation.2. Clogged drain hole.6. Improper ring side or end gapclearance.7. Plugged packing vent system.8. Scored piston rod.9. Excessive piston rod run-out.Piston Rod OilScraper Leaks1. Worn scraper rings.2. Scrapers incorrectly assembled.3. Worn/scored rod.4. Improper fit of rings to rod/sideclearance.Courtesy of Ingersoll-Rand Co.13-20

Standard 618, Reciprocating Compressors for General RefineryServices.As pressure vessels, all pulsation chambers (volume bottlesand dampeners) are generally built to Section VIII of ASMECode and suitable for applicable cylinder relief valve set pressure.Suction pulsation chambers are often designed for the samepressure as the discharge units, or for a minimum of 2/3 of thedesign discharge pressure.TroubleshootingMinor troubles can normally be expected at various timesduring routine operation of the compressor. These troubles aremost often traced to dirt, liquid, and maladjustment, or to operatingpersonnel being unfamiliar with functions of the variousmachine parts and systems. Difficulties of this type canusually be corrected by cleaning, proper adjustment, eliminationof an adverse condition, or quick replacement of a relativelyminor part.Major trouble can usually be traced to long periods of operationwith unsuitable coolant or lubrication, careless operationand routine maintenance, or the use of the machine on a servicefor which it was not intended.A defective inlet valve can generally be found by feeling thevalve cover. It will be much warmer than normal. Dischargevalve leakage is not as easy to detect since the discharge isalways hot. Experienced operators of water-cooled units canusually tell by feel if a particular valve is leaking. The bestindication of discharge valve trouble is the discharge temperature.This will rise, sometimes rapidly, when a valve is in poorcondition or breaks. This is one very good reason for keepinga record of the discharge temperature from each cylinder.Recording of the interstage pressure on multistage units isvaluable because any variation, when operating at a givenload point, indicates trouble in one or the other of the twostages. If the pressure drops, the trouble is in the low pressurecylinder. If it rises, the problem is in the high pressure cylinder.Troubleshooting is largely a matter of elimination based ona thorough knowledge of the interrelated functions of the variousparts and the effects of adverse conditions. A complete listof possible troubles with their causes and corrections is impractical,but the following list of the more frequently encounteredtroubles and their causes is offered as a guide(Fig. 13-27).CENTRIFUGAL COMPRESSORSThis section is intended to supply information sufficientlyaccurate to determine whether a centrifugal compressorshould be considered for a specific job. The secondary objectiveis to present information for evaluating compressor performance.Fig. 13-28 gives an approximate idea of the flow range thata centrifugal compressor will handle. A multi-wheel (multistage)centrifugal compressor is normally considered for inletvolumes between 500 and 200,000 inlet acfm. A single-wheel(single stage) compressor would normally have application between100 and 150,000 inlet acfm. A multiwheel compressorcan be thought of as a series of single wheel compressors containedin a single casing.Most centrifugal compressors operate at speeds of3,000 rpm or higher, a limiting factor being impeller stressconsiderations as well as velocity limitation of 0.8 to0.85 Mach number at the impeller tip and eye. Recent advancesin machine design have resulted in production of someunits running at speeds in excess of 40,000 rpm.Centrifugal compressors are usually driven by electric motors,steam or gas turbines (with or without speed-increasinggears), or turboexpanders.There is an overlap of centrifugal and reciprocating compressorson the low end of the flow range, see Fig. 13-3. On thehigher end of the flow range an overlap with the axial compressorexists. The extent of this overlap depends on a numberof things. Before a technical decision could be reached as tothe type of compressor that would be installed, the service,operational requirements, and economics would have to beconsidered.Performance CalculationsThe operating characteristics must be determined before anevaluation of compressor suitability for the application can bemade. Fig. 13-29 gives a rough comparison of the characteristicsof the axial, centrifugal, and reciprocating compressor.The centrifugal compressor approximates the constanthead-variable volume machine, while the reciprocating is aconstant volume-variable head machine. The axial compressor,which is a low head, high flow machine, falls somewherein between. A compressor is a part of the system, and its performanceis dictated by the system resistance. The desired systemcapability or objective must be determined before acompressor can be selected.Fig. 13-30 is a typical performance map which shows thebasic shape of performance curves for a variable-speed centrifugalcompressor. The curves are affected by many variables,such as desired compression ratio, type of gas, numberof wheels, sizing of compressor, etc.With variable speed, the centrifugal compressor can deliverconstant capacity at variable pressure, variable capacity atconstant pressure, or a combination variable capacity andvariable pressure.Basically the performance of the centrifugal compressor, atspeeds other than design, is such that the capacity will varydirectly as the speed, the head developed as the square of thespeed, and the required horsepower as the cube of the speed.As the speed deviates from the design speed, the error of theserules, known as the affinity laws, or fan laws, increases. Thefan laws only apply to single stages or multi-stages with verylow compression ratios or very low Mach numbers.Fan Laws:Q ∝ N; i.e., Q 110N 110= Q 100N 100= Q 90N 90Eq 13-22H ∝ N 2 ; i.e.,H 110(N 110 ) 2 = H 100(N 100 ) 2 = H 90(N 90 ) 2 Eq 13-23Bhp ∝ N 3 ; i.e., Bhp 110(N 110 ) 3 = Bhp 100(N 100 ) 3 = Bhp 90(N 90 ) 3 Eq 13-24By varying speed, the centrifugal compressor will meet anyload and pressure condition demanded by the process systemwithin the operating limits of the compressor and the driver.It normally accomplishes this as efficiently as possible, since13-21

Nominal flowrange(inlet acfm)FIG. 13-28Centrifugal Compressor Flow RangeAveragepolytropicefficiencyonly the head required by the process is developed by the compressor.This compares to the essentially constant head developedby the constant speed compressor.Fig. 13-30 depicts typical performance curves with a smallcompression ratio. The system resistance has been superimposedon the chart: Line A represents typical system resistanceof a closed system, such as a refrigeration unit wherethere is a relatively constant discharge pressure. Line B is anopen-end system, such as pipeline application where pressureincreases with capacity.Fig. 13-31 shows a higher compression ratio. The range ofstable operation is reduced because of the larger compressionratio. This is indicated by the surge line in Fig. 13-31 beingfurther to the right than in Fig. 13-30.Estimating PerformanceAverageisentropicefficiencySpeed todevelop10,000 fthead/wheel100-500 0.63 0.60 20,500500-7,50 0.74 0.70 10,5007,500-20,000 0.77 0.73 8,20020,000-33,000 0.77 0.73 6,50033,000-55,000 0.77 0.73 4,90055,000-80,000 0.77 0.73 4,30080,000-115,000 0.77 0.73 3,600115,000-145,000 0.77 0.73 2,800145,000-200,000 0.77 0.73 2,500FIG. 13-29Compressor HeadFigs. 13-32 through 13-39 may be used for estimating compressorperformance. These curves are suitable for estimatingonly and are not intended to take the place of a "wheel-bywheel"selection by the compressor manufacturer, nor shouldthe curves be used to calculate performance using field datain an attempt to determine a variance from predicted performancebased on manufacturer’s data. Fig. 13-32 is used to convertscfm to icfm. All centrifugal compressors are based onflows that are converted to inlet or actual cubic feet per minute.This is done because the centrifugal wheel is sensitive toinlet volume, compression ratio (i.e., head), and specific speed.Fig. 13-33 is a useful curve to find inlet (actual) cfm whenthe weight flow in lb/min is known. Actual cfm and inlet cfmboth denote the gas at suction conditions. These terms areoften used interchangeably. This curve can be used in reverseto determine mass flow.Fig. 13-34 is used to determine the approximate dischargetemperature that is produced by the compression ratio. Dischargetemperatures above the 400°F range should be checkedsince mechanical problems as well as safety problems mayexist. This curve includes compressor efficiencies in the rangeof 60 to 75%.Example 13-4 — Given: r = 10.0; Q 1 = 10,000 icfmFind: Discharge temperaturek = 1.15; t 1 = 0°FAnswer: t 2 = 230°F (approximately) from Fig. 13-34.Fig. 13-36 gives the approximate horsepower required forthe compression. It includes compressor efficiencies in therange of 60 to 70%.Example 13-5 — Given: Weight flow, w, = 1,000 lb/minFind: Horsepowerhead = 70,000 ft-lb/lbAnswer: Ghp = 3,000 from Fig. 13-36.Fig. 13-39 predicts the approximate number of compressorwheels required to produce the head. If the number of wheelsis not a whole number, use the next highest number.Calculating PerformanceWhen more accurate information is required for compressorhead, gas horsepower, and discharge temperature, the equationsin this section should be used. This method applies to agas mixture for which a P-H diagram chart is not available.To calculate the properties of the gas, see Figs. 13-6 and 13-7.All values for pressure and temperature in these calculationprocedures are the absolute values. Unless otherwise specified,volumes of flow in this section are actual volumes.To calculate the inlet volume:Q = (w) (1,545) (T 1) (Z 1 )Eq 13-25(MW) (P 1 ) (144)If we assume the compression to be isentropic (reversibleadiabatic, constant entropy), then:(k − 1)/kZRT ⎡⎛P 2 ⎞H is =⎢⎜MW (k − 1)/k ⎣⎝P ⎟⎠ − 1 ⎤ ⎥ Eq 13-261 ⎦Since these calculations will not be wheel-by-wheel, thehead will be calculated across the entire machine. For this, usethe average compressibility factor:Z 1 + Z 2Z avg =2The heat capacity ratio, k, is normally determined at theaverage suction and discharge temperature (see Figs. 13-7and 13-8).13-22

FIG. 13-30Compressor Performance, Low Compression RatioFIG. 13-31Compressor Performance, Higher Compression RatioIsentropic CalculationTo calculate the head:H is =Z avg RT 1MW (k − 1)/kwhich can also be written in the form:(k − 1)/k⎡⎛P 2 ⎞⎢⎜⎣⎝P ⎟⎠ − 1 ⎤ ⎥ Eq 13-27a1 ⎦H is = 1545(k − 1)/kZ avg T 1 ⎡⎛P 2 ⎞MW (k − 1)/k⎢⎜⎣⎝P ⎟⎠ − 1 ⎤ ⎥ Eq 13-27b1 ⎦The gas horsepower can now be calculated from:Ghp =(w) (H is )(η is ) (33,000)Eq 13-28The approximate theoretical discharge temperature can becalculated from:∆T ideal = T 1 ⎡ ⎢⎣⎛⎜⎝P 2P 1⎞⎟⎠(k − 1)/k− 1 ⎤ ⎥⎦Eq 13-29T 2 = T 1 + ∆T ideal Eq 13-30The actual discharge temperature can be approximated:∆T actual = T 1(k − 1)/k⎡⎛P 2 ⎞⎢⎜⎣⎝P ⎟⎠ − 1 ⎤ ⎥1 ⎦Eq 13-31η isT 2 = T 1 + ∆T actual Eq 13-32Polytropic CalculationSometimes compressor manufacturers use a polytropic pathinstead of isentropic. Polytropic efficiency is defined by:n(n − 1) = ⎡ k ⎤⎢⎣(k − 1)⎥ η p Eq 13-33⎦(See Fig. 13-37 for conversion of isentropic efficiency topolytropic efficiency.)The equations for head and gas horsepower based uponpolytropic compression are:(n − 1)/nZ avg RT 1 ⎡⎛P 2 ⎞H p =MW (n − 1)/n⎢⎜⎣⎝P ⎟⎠ − 1 ⎤ ⎥ Eq 13-34a1 ⎦which also can be written in the form:13-23

FIG. 13-32Actual Inlet Cubic Feet Per MinuteZ = 113-24

FIG. 13-33Actual Inlet Cubic Feet Per MinuteZ = 113-25

FIG. 13-34Discharge TemperatureZ = 113-26

FIG. 13-35HeadZ = 113-27

FIG. 13-36Horsepower DeterminationFIG. 13-37Efficiency Conversion13-28

FIG. 13-38Mechanical LossesFIG. 13-39Wheels RequiredCasing SizeMax Flow(inlet acfm)Nominal Speed(rpm)1 7,500 10,5002 20,000 8,2003 33,000 6,4004 55,000 4,9005 115,000 3,6006 150,000 2,800H p = 1545(n − 1)/nZ avg T 1 ⎡⎛P 2 ⎞MW⎢⎜(n − 1)/n ⎣⎝P ⎟⎠ − 1 ⎤ ⎥ Eq 13-34b1 ⎦Ghp =a. Bearing horsepower losses807060504030Bearing loss, hp2010Casing size, #671 2 3 4 5 6 8 10 20Shaft speed, N, thousand rpm(w) (H p )(η p ) (33,000)Polytropic and isentropic head are related byEq 13-35H p = H is η pη isEq 13-36Mechanical Losses#5b. Oil-seal horsepower losses70605040Oil-seal loss, hp302010Casing size, #6After the gas horsepower has been determined by eithermethod, horsepower losses due to friction in bearings, seals,and speed increasing gears must be added.Fig. 13-38 shows losses related to the shaft speed and casingsize for conventional multistage units.#5#4#4#3#3#2 and #1#2 and #171 2 3 4 5 6 8 10 20Shaft speed, N, thousand rpmCourtesy Chemical Engineering MagazineBearings and seal losses can also be roughly computed fromScheel’s equation:Mechanical losses = (Ghp) 0.4 Eq 13-37To calculate the brake horsepower:Bhp = Ghp + mechanical lossesEq 13-38Compressor SpeedThe basic equation for estimating the speed of a centrifugalcompressor is:N = (N nominal ) √⎺⎺⎺⎺⎺⎺⎺H totalEq 13-39(s) (H max/stage)where the number of wheels is determined from Fig. 13-39.Nominal speeds to develop 10,000 feet of head/stage can bedetermined from Fig. 13-28. However, to calculate the maximumhead per stage, the following equation based on molecularweight (or more accurately, density) can be used.H max/stage = 15,000 − 1,500 (MW) 0.35 Eq 13-40This equation will give a head of 10,000 ft for a gas whenMW = 30 and 11,000 ft when MW = 16.P-H DiagramWhen a P-H diagram is available for the gas to be compressed,the following procedure should be used. Fig. 13-40represents a section of a typical P-H diagram.For the given inlet conditions, the enthalpy can be shown aspoint 1 on the P-H diagram. Starting from Point 1 follow theline of constant entropy to the required discharge pressure(P 2 ), locating the isentropic discharge state point (2 is ). Withthese two points located the differential isentropic enthalpycan be calculated from the following equation:∆ h is = h 2is − h 1 Eq 13-41To convert to isentropic head, the equation is:H is = ∆ h is (778 ft • lb/Btu) Eq 13-42To find the discharge enthalpy:13-29

h 2 = ∆h isη is+ h 1 Eq 13-43The actual discharge temperature can now be obtained fromthe P-H diagram. The gas horsepower can be calculated usingEq 13-28 and Eq 13-35.From Fig. 13-39 and Eqs. 13-39 and 13-40, the speed andnumber of wheels can be estimated.To convert to polytropic head it will be necessary to assumea polytropic efficiency. See Fig. 13-28 for an efficiency correspondingto the inlet flow. Fig. 13-37 will give a correspondingadiabatic efficiency. The polytropic head may now be determinedfrom Eq. 13-36.When a P-H diagram is available, it is the fastest and mostaccurate method of determining compressor horsepower anddischarge temperature.Centrifugal Refrigeration CompressorsCompression ratio per wheel will vary on the order of 1.5 to2.75 per wheel depending on the refrigerant and speed.Due to the ease of applying external side loads to centrifugalmachines, it is quite common to flash refrigerant from the condenseren route to the evaporators and/or to accept side loadsfrom product being cooled by refrigerant at higher pressuresthan the lowest evaporator level.Since side-loading is the practice rather than the exception,it is common to let the centrifugal compressor manufacturerobtain the desired performance characteristics from the followingdata: evaporator temperature levels; refrigerationloads required in MMBtu/hr; heat rejection medium (air orwater); and type of driver.Refrigeration compressors are also discussed in Section 14.Flow LimitsFIG. 13-40P-H Diagram ConstructionGENERALTwo conditions associated with centrifugal compressors aresurge (pumping) and stone-wall (choked flow).At some point on the compressor’s operating curve there existsa condition of minimum flow/maximum head where thedeveloped head is insufficient to overcome the system resistance.This is the surge point. When the compressor reachesthis point, the gas in the discharge piping back-flows into thecompressor. Without discharge flow, discharge pressure dropsuntil it is within the compressor’s capability, only to repeat thecycle.The repeated pressure oscillations at the surge point shouldbe avoided since it can be detrimental to the compressor. Surgingcan cause the compressor to overheat to the point the maximumallowable temperature of the unit is exceeded. Also,surging can cause damage to the thrust bearing due to therotor shifting back and forth from the active to the inactiveside."Stonewall" or choked flow occurs when sonic velocity isreached at any point in the compressor. When this point isreached for a given gas, the flow through the compressor cannotbe increased further without internal modifications.Interstage CoolingMultistage compressors rely on intercooling whenever theinlet temperature of the gas and the required compression ratioare such that the discharge temperature of the gas exceedsabout 300°F.There are certain processes that require a controlled dischargetemperature. For example, the compression of gasessuch as oxygen, chlorine, and acetylene requires that the temperaturebe maintained below 200°F.The thermal stress within the horizontal bolted joint is thegoverning design limitation in a horizontally split compressorcase. The vertically split barrel-type case, however, is free fromthe thermal stress complication.Substantial power economy can be gained by precooling thegas before it enters the interstage impellers. Performance calculationsindicate that the head and the horsepower are directlyproportional to the absolute gas temperature at eachimpeller.The gas may be cooled within the casing or in external heatexchangers.Two methods of cooling within the casing are used — watercooled diaphragms between successive stages and direct liquidinjection into the gas.Diaphragm cooling systems include high-velocity water circulationthrough cast jackets in the diffuser diaphragms. Thediaphragm coolers are usually connected in series.Liquid injection cooling is the least costly means of controllingdischarge temperatures. It involves injecting and atomizinga jet of water or a compatible liquid into the returnchannels. In refrigeration units, liquid refrigerant is frequentlyused for this purpose. Injected liquid also functions asa solvent in washing the impellers free of deposits. Nevertheless,the hazards of corrosion, erosion, and flooding presentcertain problems resulting in possible replacement of the compressorrotor.External intercoolers are commonly used as the most effectivemeans of controlling discharge temperatures. The gas isdischarged from the compressor casing after one or morestages of compression and, after being cooled, is returned tothe next stage or series of stages for further compression.13-30

Intercoolers usually are mounted separately. When thereare two or more compressor casings installed in series, individualmachines may or may not be cooled or have intercoolers.In some cases, it may be advantageous to use an externalcooler to precool gas ahead of the first wheel.Journal and Thrust BearingsRadial journal bearings are designed to handle high speedsand heavy loads and incorporate force-feed lubrication. Theyare self-aligning, straight sleeve, multi-lobe sleeve, or tiltingpad type, each sized for good damping characteristics and highstability.Tilting pad bearings have an advantage over the sleeve typeas they eliminate oil whip or half-speed oil whirl which cancause severe vibrations.Bearing sleeves or pads are fitted with replaceable steelbackedbabbitted shells or liners.Axial thrust bearings are bidirectional, double faced, pivoted-shoetype designed for equal thrust capacity in both directionsand arranged for force-feed lubrication on each side.Thrust bearings are sized for continuous operation at maximumdifferential pressure including surge thrust loads, axialforces transmitted from the flexible coupling and electric motorthrust.On units where the thrust forces are low, a tapered landthrust bearing may be used but must be selected for properrotation direction. At times a combination of pivoted-shoe andtapered land is recommended.Compressor designs with impellers arranged in one directionusually have a balance drum (piston) mounted on the dischargeend of the shaft to minimize axial loads on the thrustbearing.Fig. 13-41 illustrates typical journal and thrust bearingsgenerally used in horizontally and vertically split casings.Bearing supports are cast integral with, or bolted to, the casewith isolated bearing chambers to prevent lubricating oil leakageinto the gas system or contamination of the oil or the gas.Bearing housings are horizontally split and readily accessiblefor inspection and maintenance. Provisions are made to accommodatepick-ups and sensors for vibration and temperaturemonitoring.Shaft SealsShaft seals are provided on all centrifugal compressors tolimit, or completely eliminate, gas leakage along the shaftwhere it passes through the casing.With the wide range of temperature, pressure, speed, andoperating conditions encountered by compressors, there canbe no one universal seal, or seal system, to handle all applications.Basically, the designs of seals available are: labyrinth (gas),restrictive ring (oil or gas), liquid film (oil), and mechanical(contact) (oil or gas).A mechanical (contact) seal, Fig. 13-42, has the basic elementssimilar to the liquid film seal. The significant differenceis that clearances in this seal are reduced to zero. The sealoperates with oil pressure 35 to 50 psi above internal gas pressureas opposed to 5 psi in the liquid film seal.FIG. 13-41Journal and Thrust Bearing AssemblyThe mechanical (contact) seal can be applied to most gases,but finds its widest use on clean, heavier hydrocarbon gases,refrigerant gases, etc.A mechanical gas seal uses the process gas as working fluidto eliminate the seal oil system.The liquid film seal, Figs. 13-43 and 13-44, was also developedfor the severe conditions of service but requires higheroil circulation rate than the mechanical (contact) type.The seal consists of two sleeves which run at close clearanceto the shaft with a liquid injected between the sleeves to flowto the seal extremities. The sleeves are lined with babbitt ora similar non-galling material which is compatible with theproperties of the compressed gas and the sealing liquid.The sealing liquid, usually a lubricating oil, is introducedbetween the two rings at a controlled differential pressure ofabout 5 psi above the internal gas pressure, presenting a barrierto direct passage of gas along the shaft. This fluid alsoperforms the very important functions of lubricating thesleeves and removing heat from the seal area.Lubrication and Seal-oil SystemsOn all centrifugal compressors that have force-feed lubricatedbearings, a lubrication oil system is required. When oilfilmor mechanical (contact) seals are used, a pressurizedseal-oil system must be provided.Each system is designed for continuous operation with allthe elements (oil reservoir, pumps with drivers, coolers, filters,pressure gauges, control valves, etc.) piped and mounted on aflat steel fabricated base plate located adjacent to the compressor.The compressor manufacturer normally supplies both systemsin order to have overall unit responsibility.13-31

FIG. 13-42Mechanical (Contact) Shaft SealFIG. 13-43Liquid Film Shaft Seal with Pumping BushingDepending on the application, lubrication and seal-oil systemsmay be furnished as combined into one system, or as onelubrication system having booster pumps to increase the pressureof only the seal oil to the required sealing level. In serviceinvolving heavily contaminated gases, separate lube-oil andseal-oil systems should be used.The lubrication system may supply oil to both compressorand driver bearings (including gear) couplings (if continuouslylubricated), as well as turbine governor, trip and throttlevalve, and hydraulic control system.FIG. 13-44Liquid Film Shaft Seal with Cylindrical BushingA single lubricant shall be used in all system equipment,usually an oil, having approximate viscosities of 150 SayboltUniversal seconds (SUS) at 100°F and 43 SUS at 210°F.In addition to all the elements of a common pressurized lubricationsystem, the seal oil system requires a collection systemfor the oil. Depending on the gas composition, a degassingtank may be installed in the seal oil trap return line to removethe oil-entrained gas prior to return of the seal oil to the commonoil reservoir. The flow past the outer sleeve passesthrough an atmospheric drain system and is returned to thereservoir. The relatively low flow through the inner sleeve iscollected in a drain trap or continuous drainer and may bereturned to the reservoir or discarded, depending upon thedegree and type of contamination which occurred while it wasin contact with the internal gas.Compressors using only liquid film seals should be providedwith a seal-oil system which incorporates an overhead surgetank. The surge tank provides seal-oil capacity for coastdownin case of a compressor shutdown.In combined seal-oil and lube-oil systems when largeamounts of contaminants are present in the process gas, the13-32

seal-oil design may call for buffer gas injection to form a barrierbetween the compressed gas and the seal oil.Fig. 13-45 shows clean sweet buffer gas being injected intothe center of a labyrinth seal preceding the oil film seal withseal oil supplied between the two sleeves. Part of the seal oilflows across the inner sleeve and mixes with buffer gas andthen drains into the seal oil trap. The other part of the seal oilflows across the outer sleeve, mixes with the bearing lube oildrain flow, and returns to the common lube- and seal-oil reservoir.Drivers — Centrifugal compressors can be driven by awide variety of prime movers including electric motors, steamturbines, gas combustion turbines, and gas-expander turbines.Each driver has its own design parameters. A motordrive presents limitations in operation of the compressor dueto constant and low speed. The constant speed restriction isminimized by suction or discharge throttling. The low speedrestriction is corrected by introduction of a speed increasinggear. A steam turbine, on the other hand, has variable speedcapability that allows more control of the compressor capacityor discharge pressure, and its high speed permits the compressorto be directly connected to the driver. In the case of a single-shaftgas turbine, the power output is limited at a reducedspeed.Drivers are discussed in Section 15.CONTROL SYSTEMSCentrifugal compressor controls can vary from the very basicmanual recycle control to elaborate ratio controllers. Thedriver characteristics, process response, and compressor operatingrange must be determined before the right controls canbe selected.The most efficient way to match the compressor characteristicto the required output is to change speed in accordancewith the fan laws (affinity laws, see Eqs 13-22 to 13-24):N 1N 2=Q 1Q 2=√⎺⎺ ⎺H 1√⎺⎺ ⎺H 2FIG. 13-45Combined Seal-Oil and Lube-Oil System withExternal Sweet Buffer GasOne of the principal advantages of using steam or gas turbinesas drivers for compressors is that they are well suited tovariable-speed operation. With such drivers, the speed can becontrolled manually by an operator adjusting the speed governoron the turbine or, alternatively, the speed adjustmentcan be made automatically by a pneumatic or electric controllerthat changes the speed in response to a pressure or flowsignal.Pressure Control at Variable SpeedThe control system operates as follows:The pressure transmitter (PT) in Fig. 13-46 senses the processpressure. It converts this signal to a signal proportional tothe process pressure and sends it to the pressure controller(PC).The pressure controller amplifies the transmitter signal andsends a modified signal to the final element. Depending onsystem requirements the controller may require additionalcorrection factors called reset and rate.The final element in this case is speed control. This mechanismvaries the turbine-governor speed setting over a predeterminedrange.As the load decreases, pressure will rise. An increase in processpressure above the set value will cause the signal to reachthe governor and reduce the speed, maintaining the desiredsystem pressure.FIG. 13-46Pressure Control at Variable SpeedVolume Control at Variable SpeedIf the nature of the process requires constant volume delivered,then the arrangement shown in Fig. 13-47 would beused.Here, the flow transmitter (FT) senses the process flow, convertsthe signal to a signal proportional to the process flow,and sends it to the flow controller (FC).The flow controller amplifies the transmitter signal andsends a modified signal to the final element. Reset and ratecorrection factors may be needed.The final element is speed control, which is accomplished bya mechanism that varies the turbine-governor speed setting.An increase in flow over set point would cause a signal to reachthe governor and reduce the speed to maintain the desiredsystem flow.When using electric motors as constant speed drivers, thecentrifugal compressor is normally controlled by a suctionthrottling device such as butterfly valve or inlet guide vanes.Throttling the suction results in a slightly lower suction pres-13-33

sure than the machine is designed for, and thus requires ahigher total head if the discharge pressure remains constant.This can be matched to the compressor head-capacity curve,i.e., higher head at reduced flow. In throttling the inlet, thedensity of the gas is reduced, resulting in a matching of therequired weight flow to the compressor inlet-volume capabilitiesat other points on the head/capacity curve.FIG. 13-48Pressure Control at Constant SpeedFIG. 13-47Volume Control at Variable SpeedFIG. 13-49Volume Control at Constant SpeedPressure Control at Constant SpeedThe control system shown in Fig. 13-48 has the pressuresignal sensed and amplified in a similar manner as describedin the scheme for variable speed control (Fig. 13-46).The final element is a suction throttle valve (STV) that reducesthe flow of gas into the compressor.A process pressure increase over a set value would cause asignal to reach the suction throttle valve (STV) and would partiallyclose the valve in order to reduce the inlet pressure.Volume Control at Constant SpeedThe control scheme for this arrangement is shown in Fig.13-49.The flow transmitter (FT) senses the process flow using anorifice or venturi as the primary flow element (FE), convertsthis to a signal that is proportional, and sends this signal tothe flow controller (FC). The flow controller amplifies thetransmitter signal and sends a modified signal to the finalelement. Reset and derivative controller actions may be required.The final element is the compressor guide-vane mechanism.The guide vanes are adjusted by means of a positioning cylinder.This cylinder is operated by a servo-valve (SRV) that receivesa signal from the flow controller.Here, an increase in flow above the set point causes a signalto reach the final element, which will result in the closing ofthe guide vanes to decrease flow.Adjustable Inlet Guide Vanes — The use of adjustableinlet guide vanes is the most efficient method of controllinga constant speed compressor. The vanes are built into theinlet of the 1st stage, or succeeding stages, and can be controlledthrough the linkage mechanism either automatically ormanually.The vanes adjust the capacity with a minimum of efficiencyloss and increase the stable operating range at design pressure.This is accomplished by pre-rotation of the gas enteringthe impeller which reduces the head-capacity characteristicsof the machine. Fig. 13-50 illustrates the effect of such controlat various vane positions.Prior to control selection, the economics of inlet guide vanesmust be considered because of their higher initial cost, complexmechanism, maintenance, and requirement for frequentadjustment.Anti-surge ControlIt is essential that all centrifugal compressor control systemsbe designed to avoid possible operation in surge whichusually occurs below 50% to 70% of the rated flow.Compressor surge is a large pressure and volume fluctuationthat takes place when attempting to operate at a higherpressure ratio than design maximum. The surge limit line (seeFig. 13-50) can be reached from a stable operating point byeither reducing flow or decreasing suction pressures. An antisurgesystem senses conditions approaching surge, and maintainsthe unit pressure ratio below the surge limit by recyclingsome flow to the compressor suction. Care must be taken tocool this recycle stream.Volume, pressure rise, or pressure ratio may be used as controlparameters to sense an approaching surge condition. Sucha condition will be established by the characteristic curve ofthe compressor.A volume-controlled anti-surge system is shown in Fig.13-51. The flow transmitter (FT) senses the process flow usingan orifice or venturi as a primary flow element (FE). It convertsthe signal to one that is proportional to the process flowand sends it to the surge controller (SC).The surge controller compares the transmitted signal to theset-point signal. If the set-point signal is exceeded, the controlleramplifies the signal difference and sends this modifiedsignal to the final element. Reset and rate correction factorsmay be needed.13-34

FIG. 13-50Effect of Adjustable Inlet Guide Vanes onCompressor PerformanceFIG. 13-51Anti-Surge Control–Minimum VolumeFIG. 13-52Anti-Surge Control–Pressure LimitingThe final element is a surge control valve (SCV). The valvereleases pressure buildup at the discharge of the compressor.As flow decreases to less than the minimum volume setpoint,a signal will cause the surge control valve to open. Thevalve opens, as required, to keep a minimum volume flowingthrough the compressor.For a pressure-limiting anti-surge control system, see Fig.13-52. The pressure transmitter (PT) senses the process pressure.It converts this to a signal that is proportional to processpressure and sends it to the surge controller (SC).The surge controller compares the transmitted signal to theset-point signal. If the set-point signal is exceeded, the controlleramplifies the signal difference and sends this modifiedsignal to the surge control valve. Reset and rate correctionfactors may also be required. The blow-off valve relieves apressure buildup on the discharge end of the compressor.A process pressure increase over the pressure set-point signalwill cause the blowoff valve to open. The valve opens asrequired to keep the pressure limited to a minimum volumeof gas flowing through the compressor.A set-point computer may be required on multiparametercompressor-control systems where the surge set point maychange. Suction temperature, suction pressure, and speed areexamples of parameters that cause the surge point of the compressorto change.In the design of a compressor system, any changes that mayoccur from a single operating point in the overall process mustbe carefully considered. In many instances, automatic controlsare needed to maintain the highest degree of system performance.In any event, the choice of manual or automatic controlsis dictated by the operating pattern of the centrifugal compressor.The user, compressor manufacturer, and contractor shouldwork closely together to determine the minimum control systemrequired. Various phases of operation, such as start-up,shut-down, initial runs, and normal runs, should be investigatedto make certain that a workable system has been designed.Vibration Control SystemThis control system may be provided to monitor the driverbehavior at the shaft bearings for detection of excessive lateralvibration and axial movement and for protection against possiblemachinery failure through alarm and/or shutdown devices.The system may protect not only the compressor but alsothe driver, such as a steam or gas turbine, that usually runsat the same high spead as the compressor. When a speed increasingor reducing gear unit is furnished between the compressorand driver, also consider monitoring vibration at thegear shaft bearings.The main system components are: variation transducer(s),signal amplifier(s) with d-c power supply, and vibration monitorand/or analyzer.Vibration transducers fall into three categories: displacementprobe, velocity pick-up, and accelerometer.The displacement probe is most commonly used for equipmentwith high value, as it can measure shaft vibration relativeto bearing housing. Output signal from each transduceris small and, therefore, it must be amplified before beingtransmitted to a vibration monitor or analyzer.Fig. 13-53 shows a vibration severity chart for use as a guidein judging vibration levels as a warning of impending trouble.For more information on vibration monitoring systems, seeAPI Standard 670, Noncontacting Vibration and Axial PositionMonitoring System, and API Standard 678, Accelerometer-BasedVibration Monitoring System.13-35

OPERATIONAL CONSIDERATIONSRotor Dynamics and Critical SpeedsThe demand for smooth-running turbomachinery requirescareful analysis of rotor dynamics taking into account bearingperformance, flexibility, critical speed, and rotor response.Equally important is to analyze the dynamic behavior of thecompressor for sudden changes in load due to start-up, shutdown,or loss of power supply.Successful rotor design is the result of accurate calculationof critical speeds. A critical speed occurs at a condition whenthe rotor speed corresponds to a resonant frequency of the rotor-bearingsupport system. Under no circumstances shouldthe compressor be allowed to run at a critical speed for a prolongedlength of time as the rotor vibrations amplified by thiscondition can cause machinery failure.Critical Speed MapA critical speed map is one of various methods used to predictthe operational behavior of the rotor. First, the criticalspeeds for a given rotor geometry are calculated for a range ofassumed bearing support stiffness values. The result is a maplike that shown in Fig. 13-54. The bearing stiffness characteristicsare determined from the geometry of the bearing supportsystem, and cross-plotted on the critical speed map.The map depicts the values of the undamped critical speedsand how they are influenced by bearing stiffness. The intersectionsof the bearing stiffness curve and the critical speedlines represent the undamped critical speeds. The intersectionpoints generally indicate margins between the criticals andthe operating speed range.However, the use of this map is very limited because it isbased on a simplified undamped, circular synchronous analysiswith no cross-coupled or unbalance effects. It is a goodtrending tool showing a machine’s basic dynamic characteristics.It may not accurately depict peak response frequencies.The critical speed map is used extensively because it enablesdetermination of bearing or support stiffness by correlatingtest-stand data.Unbalance Response AnalysisThis method predicts rotor-bearing system resonances togreater accuracy than the critical speed map. Here, bearingsupport stiffness and damping are considered together withsynchronous vibration behavior for a selected imbalancedistribution. A computer is normally required to solve the resultingdifferential equations. Satisfactory results depend onthe accurate input of bearing stiffness and damping parameters.Several runs are usually made with various amounts andlocations of unbalance. The plot of results of a typical unbalanceresponse study is shown in Fig. 13-55. Each curve representsthe rotor behavior at a particular station or axial locationsuch as those corresponding to the midspan, bearings, andoverhangs.No rotor can be perfectly balanced and, therefore, it must berelatively insensitive to reasonable amounts of unbalance.The unbalance-response results predict the actual amplitudesthat permit calculations of the unbalance sensitivity.FIG. 13-53Vibration Severity Chart 1FIG. 13-54Undamped Critical Speed MapThis is expressed in mils of vibration amplitude per ounce-inchor gram-inch of unbalance.The peaks of the response curves represent the critical speedlocations. Fig. 13-56 shows limits of placement of criticalspeeds as specified in the API Standard 617, Centrifugal Compressorsfor General Refinery Services.13-36

FIG. 13-55Unbalance Response PlotFIG. 13-56Rotor Response PlotVibration Levels - milsRotational Speed - RPMCourtesy of Solar Turbines InternationalCritical speeds should not encroach upon operating speedranges, and the separation margin of encroachment (SM) fromall lateral modes is required to be at least:1. Twenty (20) percent over the maximum continuous speedfor rigid shaft rotor systems.2. Fifteen (15) percent below any operating speed andtwenty (20) percent above the maximum continuousspeed for flexible shaft rotor system.Field PerformanceOnce the compressor has been installed, quite often the performanceof the system is to be tested. The same informationis required for evaluation as was supplied for the initial selection.This information is essential:1. Inlet conditionsa. Flow (scfm, acfm, or lb/min)b. Gas analysisc. Pressure, psiad. Temperature, °F2. Intermediate conditionsa. Pressure at intermediate nozzles, psiab. Temperature at intermediate nozzles, °F3. Discharge conditionsa. Pressure, psiab. Temperature, °F4. Compressor speed, rpm5. Power requirement from driver if available (steam flow,amperes, etc.)If all these conditions were given except flow, it would notbe possible to determine the compressor operating point. Thesystem resistance curve imposed on the typical performanceis not always constant. Discharge temperature of the compressoris a good indication as to the operation of the compressor.However, such items as recycle (internal or external), differentgas analysis, different suction conditions, and different flowcan affect the values. The efficiency at part load or overloadwill normally be lower than at design conditions.Test procedures should be agreed upon between the manufacturerand user before a performance test is run. A test pointshould be run several times to see if the results can be duplicated.Calibrated instruments should be used to improve theaccuracy of the test data.TroubleshootingOperational troubles occurring in service may be due to avariety of causes.If the trouble cannot be traced to adverse gas flow conditionsor liquid "slugs" present in the system, Fig. 13-57 can be usedas a guide for troubleshooting frequently encountered problems.Careless operation and maintenance needs little comment.Lack of proper care of any machine is bound to result in asuccession of minor troubles eventually leading to a majorbreakdown.TurboexpandersThe use of turboexpanders in gas processing plants beganin the early sixties. By 1970, most new gas processing plantsfor ethane or propane recovery were being designed to incorporatethe particular advantages characteristic of an expanderproducing usable work. The trend in the gas processingindustry continues toward increased use of the turboexpander.Selection of a turboexpander process cycle is indicated whenone or more of the following conditions exist:1. "Free" pressure drop in the gas stream.2. Lean gas.3. High ethane recovery requirements (i.e., over 30% ethanerecovery).13-37

FIG. 13-57Probable Causes of Centrifugal Compressor TroubleTroubleProbable Cause(s)TroubleProbable Cause(s)Low DischargePressureCompressor SurgeLow Lube OilPressureShaft Misalignment1. Compressor not up to speed.2. Excessive compressor inlettemperature.3. Low inlet pressure.4. Leak in discharge piping.5. Excessive system demand fromcompressor.1. Inadequate flow through thecompressor.2. Change in system resistancedue to obstruction in thedischarge piping or impropervalve position.3. Deposit buildup on rotor ordiffusers restricting gas flow.1. Faulty lube oil pressure gaugeor switch.2. Low level in oil reservoir.3. Oil pump suction plugged.4. Leak in oil pump suction piping.5. Clogged oil strainers or filters.6. Failure of both main andauxiliary oil pumps.7. Operation at a low speedwithout the auxiliary oil pumprunning (if main oil pump isshaft-driven).8. Relief valve improperly set orstuck open.9. Leaks in the oil system.10. Incorrect pressure control valvesetting or operation.11. Bearing lube oil orifices missingor plugged.1. Piping strain.2. Warped bedplate, compressor ordriver.3. Warped foundation.4. Loose or broken foundationbolts.5. Defective grouting.High Bearing OilTemperatureNote:Lube oil temperatureleaving bearings shouldnever be permitted toexceed 180°F.Excessive VibrationNote:Vibration may betransmitted from thecoupled machine. Tolocalize vibration,disconnect coupling andoperate driver alone.This should help toindicate whether driveror driven machine iscausing vibration.Water In Lube Oil1. Inadequate or restricted flow oflube oil to bearings.2. Poor conditions of lube oil ordirt or gummy deposits inbearings.3. Inadequate cooling water flowlube oil cooler.4. Fouled lube oil cooler.5. Wiped bearing.6. High oil viscosity.7. Excessive vibration.8. Water in lube oil.9. Rough journal surface.1. Improperly assembled parts.2. Loose or broken bolting.3. Piping strain.4. Shaft misalignment.5. Worn or damaged coupling.6. Dry coupling (if continuouslylubricated type is used).7. Warped shaft caused by unevenheating or cooling.8. Damaged rotor or bent shaft.9. Unbalanced rotor or warpedshaft due to severe rubbing.10. Uneven build-up of deposits onrotor wheels, causing unbalance.11. Excessive bearing clearance.12. Loose wheel(s) (rare case).13. Operating at or near criticalspeed.14. Operating in surge region.15. Liquid "slugs" striking wheels.16. Excessive vibration of adjacentmachinery (sympatheticvibration).1. Condensation in oil reservior.2. Leak in lube oil cooler tubes ortube-sheet.4. Compact plant layout requirement.5. High utility costs.6. Flexibility of operation (i.e., easily adapted to wide variationin pressure and products).There are multiple factors in addition to the ones listedabove that affect a final process selection. If two or more of theabove conditions are coexistent, generally a turboexpanderprocess selection will be the best choice.Fig. 13-58 shows a typical low temperature turboexpanderprocess for recovering ethane and heavier hydrocarbons froma natural gas stream.Fig. 13-59 represents the pressure-temperature diagram forthis expander process. The solid curve represents the plantinlet gas. The solid line on the right is the dew point line. Ata fixed pressure and, if the temperature of the gas is to theright of this dew point line, the gas is 100 percent vapor. If thegas is cooled, liquid starts to condense when the temperaturereaches the dew point line. As cooling continues, more liquidis condensed until the bubble point line is reached — the solidline on the left. At this point, all of the gas is liquid. Additionalcooling results in colder liquid.Downstream of the gas treating facilities, the inlet gas isrepresented by point 1 on both Fig. 13-58 and 13-59. As the13-38

FIG. 13-58Example Expander Processgas is cooled by the gas/gas exchangers and demethanizer sideexchanger, its temperature moves along the dotted line topoint 2 (Fig. 13-59). At 2, the gas enters the expander inletseparator where the condensed liquid is separated from thevapor. This vapor now has its own pressure-temperature diagram,as represented by the dashed curve. At the expanderinlet, the gas is on its dew point line.As the gas flows through the expander, its pressure-temperaturepath is shown by the dashed line from point 2 topoint 3. Point 3 represents the outlet of the expander. The importanceof using the expander as a driver for a compressorcan be seen in Fig. 13-59. If the gas had been expanded withoutdoing any driver work, the expansion path would be frompoint 2 to point 4. This is called a Joule-Thomson, or constantenthalpy expansion. The outlet temperature and pressurewould be higher than that accomplished in the expander(nearly isentropic) expansion process.THERMODYNAMICSA turboexpander recovers useful work from the expansionof a gas stream. The process operates isentropically in theideal case and produces something less than the theoreticalwork in the real case. In the process of producing work, theexpander lowers the bulk stream temperature which can re-FIG. 13-59Pressure-Temperature Diagram for Expander ProcessNote that the pressure at Point 4 is not as low as that attainedby flow through the expander (Point 3). This is becauseit has been assumed for this example that, without the expanderrunning (therefore the brake compressor also not running),the process cannot restore the demethanizer overheadvapor to the residue gas pressure using the separate recompressoralone.Also, because the path to Point 4 is adiabatic without thegas doing work, the gas does not cool to as low a temperatureas the path to Point 3. That is, the path (2) to (3) is isentropicexpansion producing work and thereby cooling the gas morethan the simple isentropic expansion path.The higher temperature results in a reduction of productrecovery. The use of the expander brake compressor to boostthe residue gas pressure will allow a lower expansion pressurewithout the use of more residue compression.13-39

sult in partial liquefaction of the bulk stream. A simple schematicof an expander is given in Fig. 13-60.An example calculation of an expander operating in a naturalgas plant follows. The GPSA method is a hand calculationprocess which relies on various K-value and thermodynamicproperty charts available in the GPSA Data Book. Due to thelimits of accuracy in reading these charts, it is impossible toduplicate all of the numbers in the example problem in spiteof using each time the same step-by-step method. For all intentand purposes, the example calculation can be regarded assufficiently correct with a discrepancy of less than 3% for thefinal result of the expander horsepower.A simplified 3-component mixture will be used, as the exampleis general for all mixtures. The example involves the useof the K-value charts (see Section 25) and the calculation ofstream enthalpies and entropies.Thermodynamic properties are taken from Section 24. K-values are calculated from the data on the binary systemsmethane-ethane, and methane-propane, and the infinite dilutionternary system.FIG. 13-60Simple ExpanderGas inlet conditions (t 1 , P 1 ) to the expander are generally setby process balances. The outlet pressure P 2 from the expanderis set by process and recompressor power considerations. Fig.13-61 gives stream conditions with an example calculation.Outlet conditions for the expander must be determined bytrial-and-error calculations. Assume three different outlettemperatures, t′, t′′, t′′′ at the set outlet pressure of P 2(300 psia), from the expander. At each assumed temperatureand P 2 , calculate the equilibrium conditions of the bulk streamfrom the expander, and determine the stream bulk enthalpy(h) and entropy (s).For most natural gas applications, including this example,a limiting boundary value for the outlet temperature can beobtained from the methane P-H diagram. From this chart, determinethe ideal outlet temperature t 2 by isentropic expansionfrom inlet t 1 , P 1 to outlet P 2 . In most real gas mixtures, t 2will be somewhat warmer than this. In the example case, t 2for pure methane is –159°F. Therefore, assume t′ = –150°F,t′′ = –140°F, t′′′ = –130°F.Draw a plot of ∆h(h t1P1 – h t2P2 ) versus temperature at P 2 , ands versus outlet temperature at P 2 , using stream bulk propertiesas calculated in Fig. 13-61. This plot is illustrated inFig. 13-62.Flow: 60 MMscfdInlet Conditions:t 1 = –60°FP 1 = 900 psiaP 2 = 300 psiamols/hrMethane 6,185.8Ethane 263.5Propane 138.3Total mols/hr 6,587.6Total lbs/hr 113,261.0All enthalpies (h) are calculated from Figs. 24-3, 24-6, and 24-7,and all entropies (s) from Figs. 24-18, 24-19, and 24-20. See Fig.24-17 for example calculation of h and Fig. 24-2 for example calculationof s.Inlet Enthalpy, (h)Inlet Entropy, (s)Outlet Conditions118.134 Btu/lb32.58 Btu/(mol • °R)-150°F300 psia-140°F300 psia-130°F300 psiaK-value for methane 1.1589 1.3680 1.6023K-value for ethane 0.0813 0.0948 0.1127K-value for propane 0.00974 0.01022 0.01235Outlet liquid, lbs/hr 48,380 23,958 15,720Outlet vapor, lbs/hr 64,882 89,303 97,541Outlet liquid enthalpy (h)Btu/lb –53.202 –68.675 –77.500Outlet vapor enthalpy (h)Btu/lb 120.702 127.901 134.301Outlet liquid entropy (s)Btu/(mol • °R) 25.557 25.916 26.220Outlet vapor entropy (s)Btu/(mol • °R) 33.172 33.688 34.147Mixture h, Btu/lb 46.419 86.319 104.904Mixture s, Btu/(mol • °R) 30.198 32.383 33.366h inlet – h outlet = ∆ h, Btu/lb 71.715 31.815 13.230See Fig. 13-62FIG. 13-61Expander Example Calculation∆ h ideal = 27.688 Btu/lb∆ h actual = 0.8 efficiency x 27.688 = 22.150 Btu/lbt 2 at 80% expander efficiency = –135.2°F @ 300 psia.Flash @ Actual Outlet Conditions:K at –135.2°F300 psiaLiquidVaporMethane 1.4726 567 5,619Ethane 0.1019 156 107Propane 0.0104 129 9Total mols/hr 852 5,735Total lbs/hr 19,477 93,784Work produced = 22.150 Btu/lb x 113,261 lb/hr = 2,509,000Btu/hrHorsepower = 2,509,000 ÷ 2,545 = 986 hp13-40

From the plot of the calculated s versus outlet temperatureat P 2 (300 psia), determine the isentropic outlet temperature,using s 1 (32.58) from the inlet conditions. From this outlet temperature,read the corresponding ∆h ideal, which is the idealwork available from a 100% efficient isentropic process. Sincethe expander operates at something less than ideal conditions,an efficiency must be used to determine actual expander outletconditions. If an expander efficiency of 80% is used, the actual∆h = 0.8 x ∆h ideal = ∆h A . Using ∆h A from the plot, read anoutlet temperature of t 2 (–135.2°F). Now make an equilibriumflash calculation at conditions t 2 , P 2 (–135.2°F, 300 psia) to findthe real outlet conditions for the expander, operating at 80%efficiency.The total work produced by the expander based on the assumedefficiency can be calculated as shown in Fig. 13-61.A schematic P-H diagram path for the example expandercalculation is shown in Fig. 13-63.In many applications the loading device for the turboexpanderis a centrifugal compressor and its performance can bethe system. Generally these contaminants will form a blockageupstream of the expander, in the lower temperature exchangecircuit, or on the screen ahead of the expander.Carbon dioxide can also form as a solid in the system, particularlyat the lower temperatures of the expander outlet.Fig. 13-64 will provide a quick estimate for the possibility offormation of solid CO 2 . If operating conditions are in the methaneliquid region as shown by the insert graph, the dashedFIG. 13-63Schematic P-H Diagram for ExpanderFIG. 13-62T-h and T-s Diagramsolid-liquid phase equilibrium line is used. For other conditionsthe solid isobars define the approximate CO 2 vapor concentrationlimits.For example, consider a pressure of 300 psia. At –170°F, theinsert graph (Fig. 13-64) shows the operating conditions to bein the liquid phase region. The dashed solid-liquid phase equilibriumline indicates that 2.1 mol percent CO 2 in the liquidphase would be likely to form solids. However, at the samepressure and –150°F, conditions are in the vapor phase, and1.28 mole percent CO 2 in the vapor could lead to solids formation.This chart represents an approximation of CO 2 solidformation. Detailed calculations should be carried out if Fig.13-64 indicates operation in a marginal range.If the expander liquid is fed to the top tray of a demethanizer,the CO 2 will concentrate in the top equilibrium stages.This means that the most probable condition for solid CO 2formation may be several trays below the top of the towerrather than at expander outlet conditions. Again, if Fig. 13-64indicates marginal safety from solids formation, detailed calculationsmust be carried out.calculated as shown in this section. Shaft and bearing lossesin the order of 2% are usually deducted to calculate net powerinput to the driven end from the expander.Solids FormationIn addition to the obvious need for water removal from thegas stream to protect the expander in low temperature service,consideration must be given to the possible formation of othersolids or semi-solids in the gas stream. Amines, glycols, andcompressor lube oils in the gas stream can form blockages inMECHANICALMechanical design of the turboexpander is the business ofseveral manufacturers. Any specific information must comefrom such supplier.Of the various general turbine types available, the radialreaction turbine design is dominant in cryogenic turboexpandernatural gas plant applications. These units operate overwide ranges of inlet flow and pressure conditions, by utilizingvariable inlet guide vanes. They operate at very high rotating13-41

FIG. 13-64Approximate Solid CO 2 Formation Conditions13-42

speeds and thus are subject to the design and operating cautionscommon to similar sophisticated rotating equipment.The most common configuration is a turboexpander-compressorwhere the expander power is used to compress gas inthe process. In this case, the compressor wheel operates on thesame shaft as the expander wheel. Other applications of thepower recovery are expander-pump or expander-generatordrives. These normally require gearing to reduce the expanderspeed to that required for the driven unit.Since power recovery and refrigeration effect are primarybenefits of expander applications, rotating speeds are set tooptimize the expander efficiency. This will usually result in acompromise in the compressor end design and lower compressorefficiencies. Usual efficiencies quoted for radial type unitsare 75 to 85% for the expander and 65 to 80% for the compressor.Some areas requiring extra attention in the installation ofturboexpanders are listed below. The list is by no means comprehensive,but these items require more than the normalamount of concern in designing the installation of a turboexpanderunit for cryogenic operation.1. The expander inlet gas stream must be free of solid orliquid entrainment. Liquids are removed in a high pressureseparator vessel. An inlet screen of fine mesh is usuallyrequired for solids removal. Monitoring of the pressuredrop across this screen is recommended. Formationof solids (ice, carbon dioxide, amines, heavy oils) willoften occur here first and can be detected by an increasein pressure drop across the screen.2. Source of the seal gas, particularly during start-up, is animportant consideration. The stream must be clean, dry,sweet, and of sufficient pressure to meet the system requirements.3. Normally a quick closure shutoff valve is required on theexpander inlet. Selection of this valve and actuator typemust take into account start-up, operating, and shutdownconditions.4. Vibration detection instrumentation is useful but notmandatory. Its application is normally an owner and vendoroption and influenced by operating economics.5. Loading of the flanges by the process piping system mustbe within prescribed limits to avoid distortion of the case,resulting in bearing or wheel rubbing problems.6. Failures due to mechanical resonance have occurred inturboexpanders. Even though the manufacturer will exerthis best efforts at the manufacturing stage to avoidthis problem, in-plant operation may uncover an undesirableresonance. This must be solved in conjunctionwith the manufacturer and may involve a redesign of thewheels, bearing modifications, vane or diffuser redesign,etc.The installation of a turboexpander-compressor unit also requiresthe proper design of a lube system, instrumentation,etc., in common with other industrial rotating equipment. Itis common practice to install a turboexpander-compressorwith no special anti-surge instrumentation for the compressorunit. This is acceptable if it can be determined that the gasflow through the compressor is balanced with flow through theexpander and the two will vary simultaneously.Auxiliary SystemsBoth lubricated and non-lubricated turboexpander designsare available.Lubrication System — The lubrication system circulatescooled and filtered lube oils to the turboexpander bearingsas shown on Fig. 13-65. The principle components of thesystem are monitored on the lube console and normally consistof two electric motor-driven lube oil pumps, an oil cooler, a dualfilter valve, a bladder type with switching coastdown accumulator,and a pressurized reservoir with mist eliminator.The lube oil pumps (one stand-by) must maintain a constantflow to the radial and thrust bearings. Absence of oil, or improperfiltration, can cause bearing damage. Most manufacturersrecommend a light turbine oil (315 SSU at 100°F) forbest machine performance.The lube oil cooler is an integral part of the system to rejectheat that is generated across the bearings. It can be of a fanair cooled type or shell and tube design, water cooled. If thecooling water is scale forming, duplicate coolers (one stand-by)are recommended.Lube oil filtration is extremely important due to close tolerancesbetween bearing surfaces.The lube-oil reservoir serves as a surge tank to enhancepump suction as well as to serve as a degassing drum permittingprocess seal gas to be released from the oil. If necessary,the reservoir should be equipped with a heater to bring the oilup to temperature for a "cold" start.Seal Gas System — The seal gas system prevents loss ofprocess gas and assures protection against entry of lube oilinto process gas areas. To accomplish this, a stream of "sealgas" is injected into each labyrinth shaft seal at a pressurehigher than that of the process gas. The leaking seal gas iscollected in the oil reservoir, then returned through a mistFIG. 13-65Lube Oil Schematic13-43

eliminator to the fuel gas system, or put back into the compressorsuction end.The system for seal gas injection consists of a liquid collector,electric heater (if required), twin filters, and differential pressureregulators.FIG. 13-66Example Change in Efficiency with Flow RateIf recompression is necessary for the gas processing plant,sales gas is ideal for use as seal gas. If no recompression isprovided, a stream can be taken from the expander inlet separator,warmed and used as seal gas. A minimum seal gas temperature(about 70°F) is required to prevent oil thickening.Seal gas filtration is essential because of close clearancesprovided between the shaft and seals.Seal gas flow requirements are determined by the expandermanufacturers as a part of their performance rating.Control SystemsProcess — Control of the process streams begins withproper dehydration and filtering. Generally a final protectivescreen upstream of the expander is designed into the pipingsystem to form a protective barrier against carbon dioxide orwater freezing.As a further protection against water freezing, methanol injectionconnections are incorporated into the system upstreamof the expander.Machine — The expander speed is established by themanufacturer, given the process conditions. For maximum efficiency,the expander manufacturer determines the wheel diameterand specific speed.As plant operating conditions change, the expander speedmay change. Fig. 13-66 shows the change in efficiency as afunction of change in design flow rate.Gas entering the expander is directed by adjustable nozzlesinto the impeller. About one-half of the pressure drop acrossthe expander takes place in the nozzles, imparting kinetic energyto the gas which is converted to shaft horsepower by theexpander wheel. Pressure reductions are normally limited to3-4 ratios. Greater ratios reduce expander efficiency to theextent that 2-stage expansion may be advisable.The adjustable nozzles function as pressure control valves.A pneumatic operator takes a split range signal (3 to 9 psi) tostroke the nozzles. On increasing flow beyond the full opennozzle position, a 9 to 15 psi signal from a pressure controlleropens a bypass control valve. This valve is called the J-T(Joule-Thomson) valve.Thrust bearing force imbalance is caused by difference inpressures between the expander discharge and compressorsuction. With a differential of the order of 20 psi, the thrustloads are usually within the capabilities of the thrust bearings.At higher pressure differentials, it is essential that steps betaken to control the thrust loads against each other, therebythe net thrust load will not exceed the thrust bearing capacity.This is done by providing a force-measuring load-meter oneach thrust bearing, Fig. 13-67, and a thrust control valvewhich controls the thrust by control of pressure behind thethrust balancing drums or behind one of the seals. These twoload-meters indicate thrust bearing oil film pressure (proportionalto bearing load) and the third shows the pressure behindthe balancing drum as controlled by the valve in its vent as ameans of adjusting the thrust load.Vibration comes from an unbalanced force on one of the rotatingcomponents, or it could come from an outside sourcesuch as pipe vibration or gas pulsation.Most expanders are supplied with monitoring and shutdowndevices for shaft vibration. These devices are set to shut downthe expander before damage occurs.Lube Oil — The lube oil must be filtered. Most systemsuse a primary and secondary filtering system. Controls areprovided to ensure oil flow to bearings at proper pressure andtemperature. Two (2) lube oil pumps are furnished, the secondpump serving as a standby. The standby oil pump is controlledautomatically to cut in to provide oil pressure upon failure ofthe main pump or reduction in pressure for other reasons.Generally an oil flow bypass valve is included to permit excessflow to bypass the expander bearings and return to thereservoir.For temperature control, the oil must be cooled to prohibitheat buildup which occurs through the bearings. Also, atemperature control bypass is included in the circuit for anextra measure of control to keep the oil from getting too cool.Seal Gas — Use a suitable gas stream with filtering andpressure control to maintain proper gas pressure at the shaftseals.If the seal gas is delivered from a cold supply point (expanderinlet separator) then a means of heating the gas isnecessary.The seal gas should be introduced before the lube oil systemis started because there might be a pressure upset whichwould put enough oil into the process to cause a problem.Each of the main rotating components (radial bearings,thrust bearings, and shaft seals) can be damaged or eroded byimproper oil filtration, lack of oil flow, improper gas dehydration,and improper seal gas filtration.Shutdown — A number of conditions during the operationof expanders justify prompt shutdown to avoid seriousdamage.Some of these conditions are:• High Vibration• Low Lube Oil Flow• High Inlet Separator Level13-44

FIG. 13-67Typical Expander/Compressor Cross-Section with Thrust Balancing Schematic• High Inlet Screen Pressure Drops• High Thrust• High Lube Oil Temperature• Low Lube Oil Pressure• High SpeedTwo primary actions of a shutdown signal are to block gasflow to the expander and the compressor. This is accomplishedby actuating quick acting shutdown valves at the expanderinlet and outlet and the compressor inlet. Simultaneously, apressurized bladder supplies oil to the bearings during theexpander coast down. The expander bypass valve (J-T) opensautomatically and is positioned by the split-range pressurecontrolled to keep the plant on-line in the J-T mode.Field Performance — Field measurements can bemade to check efficiencies and horsepower of the expander.The process of calculations is just the reverse of selecting amachine performance.Knowing the gas composition, mass flow (lbs/hr), inlet andoutlet conditions (pressure, temperature) for the expander, theactual difference in enthalpy can be determined for each unit.Thus: ∆h actual = h t2 P 2− h t1 P 1η =∆h actual∆h ideal∆h actual • lbs/hrHP actual =2,545The above power calculated for the expander is then reconciledto the calculated power consumed by the brake compressor.The brake compressor horsepower is calculated from itsflow, suction and discharge conditions, gas composition, andan assumed efficiency. In practice the application of this techniqueis usually difficult due to lack of precise data. Processcalculations based upon overall plant balance data may yieldinformation to aid in the analysis of equipment efficiencies.REFERENCE1. ISR MECHANALYSIS, INC., 6150 Huntley Road, Columbus,Ohio 43229.BIBLIOGRAPHY1. American Petroleum Institute Standards:API 614 - "Lubrication, Shaft-Sealing and Control Oil Systemsfor Special-Purpose Applications".API 617 - "Centrifugal Compressors for General Refinery Services".API 618 - "Reciprocating Compressors for General Refinery Services".API 670 - "Non-Contacting Vibration and Axial Position MonitoringSystem".API 678 - "Accelerometer Based Vibration Monitoring Systems".2. Bergmann, D./Mafi, S., "Selection Guide for Expansion Turbines,"Hydrocarbon Processing, Aug. 1979.3. Brown, R. N., "Control Systems for Centrifugal Gas Compressors,"Chemical Engineering, Feb. 1964.4. Criqui, A. F., "Rotor Dynamics of Centrifugal Compressors," SolarTurbines International, San Jose, California.13-45

5. Gibbs, C. W., "Compressed Air and Gas Data," Ingersoll Rand Co.6. Neerken, R. F. "Compressor Selection for the Process Industries,"Chemical Engineering, Jan. 1975.7. Perry, R. H./Chilton, C. H., "Chemical Engineers Handbook,"Fifth Edition, Section 6, McGraw-Hill Book Co., Inc., New York,New York.8. Reid, C. P., "Application of Transducers to Rotating MachineryMonitoring and Analysis," Noise Control and Vibration Reduction,Jan. 1975.9. Scheel, L. F., "Gas and Air Compression Machinery," McGraw-Hill Book Co., Inc., New York, New York.10. Swearingen, J. S., "Turboexpanders and Expansion Processes forIndustrial Gas," Rotoflow Corp., Los Angeles, California.13-46