Determining Compressor Acceptability

8/14/2019 Determining Compressor Acceptability

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Note: The source of the technical material in this volume is the ProfessionalEngineering Development Program (PEDP) of Engineering Services.

Warning: The material contained in this document was developed for SaudiAramco and is intended for the exclusive use of Saudi Aramcos employees.Any material contained in this document which is not already in the publicdomain may not be copied, reproduced, sold, given, or disclosed to thirdparties, or otherwise used in whole, or in part, without the written permissionof the Vice President, Engineering Services, Saudi Aramco.

Chapter : Mechanical For additional information on this subject, contactFile Reference: MEX-212.04 PEDD Coordinator on 874-6556

Engineering Encyclopedia

Saudi Aramco DeskTop Standards

DETERMINING COMPRESSOR ACCEPTABILITY


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Engineering Encyclopedia Compressors

Determining Compressor Acceptability

Saudi Aramco DeskTop Standards i

Section Page

INFORMATION............................................................................................................... 4

INTRODUCTION............................................................................................................. 4

TEST METHODOLOGIES FOR ACCEPTABILITY OF DYNAMIC AND POSITIVE-DISPLACEMENT COMPRESSORS............................................................................... 5

Hydrostatic Test.................................................................................................... 5

Helium Leak Test.................................................................................................. 6

Mechanical Running Test (Including Rotor Dynamics) ......................................... 7

Gas Leakage Test .............................................................................................. 18

Performance Test ............................................................................................... 19

String Test .......................................................................................................... 21

Post-Test Inspection........................................................................................... 22

DETERMINING DYNAMIC COMPRESSOR ACCEPTABILITY.................................... 23

Acceptability Criteria (31-SAMSS-001)............................................................... 26

Calculating Inlet Flow ......................................................................................... 26

Polytropic Calculations ....................................................................................... 28

Calculating Pressure Ratio from Head ............................................................... 30

Calculating Horsepower and Efficiency .............................................................. 31

Use of Fan Laws to Find the Operating Point atDifference Tip Speeds........................................................................................ 35

DETERMINING POSITIVE-DISPLACEMENT COMPRESSOR ACCEPTABILITY....... 38

Acceptability Criteria (31-SAMSS-002/31-SAMSS-003)..................................... 38

Calculating Capacity........................................................................................... 39

Volumetric Efficiency................................................................................ 40

Cylinder Displacement............................................................................. 43

Percent Clearance................................................................................... 44

Calculating Discharge Temperature ................................................................... 44

Calculating Power............................................................................................... 45

WORK AIDS.................................................................................................................. 47


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Saudi Aramco DeskTop Standards ii

WORK AID 1: RESOURCES USED TO DETERMINE DYNAMIC COMPRESSORACCEPTABILITY .......................................................................................................... 47

Work Aid 1A: Calculation Procedures................................................................ 47

Work Aid 1B: Pertinent Data.............................................................................. 52

Nomenclature .......................................................................................... 52

Charts for Determining Compressor Performance Characteristics .......... 53

WORK AID 2: RESOURCES USED TO DETERMINE POSITIVE-DISPLACEMENTCOMPRESSOR ACCEPTABILITY................................................................................ 55

Work Aid 2A: Calculation Procedures................................................................ 55

Work Aid 2B: Pertinent Data.............................................................................. 58

Nomenclature .......................................................................................... 58

GLOSSARY .................................................................................................................. 60


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Saudi Aramco DeskTop Standards iii

LIST OF FIGURES

Figure 1. Typical Rotor Response Plot......................................................................... 12

Figure 2. Typical, Multi-Stage, Centrifugal Compressor Characteristic Curve.............. 24

Figure 3. Typical Axial Compressor Characteristic Curve............................................ 25

Figure 4. Adiabatic Versus Polytropic Process............................................................. 34

Figure 5. Head Curve................................................................................................... 37

Figure 6. Horsepower Curve ........................................................................................ 37Figure 8. Compressibility Factors at Low Reduced Pressure....................................... 54

Figure 9. Loss Correction Factor for Reciprocating Compressor ................................. 59

LIST OF TABLES

Table 1. Critical Constants of Gases............................................................................ 53


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Saudi Aramco DeskTop Standards 4

INFORMATION

INTRODUCTION

Gas compressor inspection and testing for acceptability areperformed as indicated on the compressor data sheets and thereferenced Saudi Aramco Form 175 based on the compressortype and the associated auxiliary equipment. The inspectionrequirements for gas compressors will vary with the compressortype and application. The Engineer must become familiar withthe requirements and criteria used for the acceptance of a gascompressor. This module provides background information onthe testing and the inspection requirements, the methods, andthe Gas compressor inspection and testing for acceptabilitycriteria for dynamic and positive-displacement compressors.


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TEST METHODOLOGIES FOR ACCEPTABILITY OF DYNAMIC ANDPOSITIVE-DISPLACEMENT COMPRESSORS

All compressor inspections and tests are to be within theguidelines and conditions that are set forth in the applicableSaudi Aramco Engineering Standard (SAES-K-402 forcentrifugal compressors and SAES-K-403 for reciprocatingcompressors). These tests and inspections include thefollowing:

Hydrostatic Test

Helium Leak Test

Mechanical Running Test

Gas Leakage Test

Performance Test

String Test

Post-Test Inspection

Before the above tests are conducted, a visual inspection of thecompressor is performed in accordance with the applicableSaudi Aramco Engineering Standards and API Standards forthe compressor to be tested.

Hydrostatic Test

Hydrostatic tests are performed by the vendor, and they do notrequire visual inspection or witnessing by a Saudi Aramcorepresentative. The vendor is required to provide Saudi Aramcowith certificates and data for the hydrostatic test results forevaluation.

Pressure-containing parts (including auxiliaries) must behydrostatically tested with liquid at a minimum of 1-1/2 (150%)times the maximum allowable working pressure but at not lessthan 20 psig for all components of a reciprocating compressor.The only exceptions to the minimum hydrostatic test pressureare the cylinder cooling jackets and packing cases, which havea minimum pressure of 115 psig. The test liquid must be at ahigher temperature than the nil-ductility transition temperature of


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the material that is being tested. The tests must be maintainedfor a sufficient period to allow a complete examination of theparts that are under pressure. The hydrostatic test will beconsidered satisfactory when neither leaks nor seepage through

the casing or casing joint is observed for a minimum of 30minutes. Large, heavy castings may require a longer testingperiod. For example, SAES-K-403 (for reciprocatingcompressors) requires that test pressure for critical items, suchas large cast cylinders, must be maintained for four hours.Seepage past internal closures that are required for testing ofsegmented cases and the operation of a test pump to maintainpressure are acceptable.

The chloride content of the liquids that are used to testaustenitic stainless steel materials in centrifugal compressors

must not exceed 50 parts per million. To prevent the depositionof chlorides that is caused by evaporative drying, all residualliquid must be removed from the tested parts at the conclusionof the test.

If the part to be tested is to operate at a temperature at whichthe strength of a material is below the strength of that materialat room temperature, the hydrostatic test pressure will bemultiplied by a factor. The factor is obtained through division ofthe allowable working stress for the material at roomtemperature by the allowable working stress for the material at

operating temperatures. The stress values that are used willconform to those values that are given in ASME B31.3 forpiping. For compressor casings and pressure vessels, thestress values must conform to those values that are given inSection VIII, Division 1 or 2, as applicable, of the ASME Code.The pressure that is obtained will then be the minimum pressureat which the hydrostatic test must be performed. The datasheets must list the actual hydrostatic test pressures.

Helium Leak TestHelium leak tests are performed by the vendor on rotary andcentrifugal compressors that are used for hydrogen service.The helium leak test must be witnessed by a Saudi Aramcorepresentative. Test documentation and data must besubmitted to Saudi Aramco for review. SAES-K-402 requires ahelium leak test to be performed on the compressor casing ofany centrifugal compressor that is in hydrogen service. API


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Standard 618 (for reciprocating compressors) requires a heliumleak test to be performed on all pressure-retaining parts, suchas cylinders and volume pockets, for all compressors thathandle gases with a molecular weight of 12 or less, or for gases

that contain more than 0.1 mol percent hydrogen sulfide. Thehelium leak test is to be performed after the hydrostatic test.The compressor casing for centrifugal compressors and thepressure-retaining parts for reciprocating compressors aretested for gas leakage with helium at the maximum allowableworking pressure. The test can be conducted in the followingtwo ways:

The casing or pressure-retaining parts are pressurized withhelium to the maximum allowable working pressure. Thecomponents are then submerged in water. The maximum

allowable working pressure must be maintained for a minimumof 30 minutes; no bubbles are permitted (zero leakage).

The casing or pressure-retaining parts are pressurized withhelium to the maximum allowable working pressure. Thepressure is maintained for a minimum of 30 minutes. Anonsubmerged soap-bubble test is performed on the casing of acentrifugal compressor. Leak detection is accomplishedthrough use of a helium probe for the pressure-retaining parts ofa reciprocating compressor. Zero leakage is required.

Mechanical Running Test (Including Rotor Dynamics)

The following discussion of test methods and requirements isderived from applicable sections of API Standard 617 and APIStandard 618. Test procedures and acceptance criteria will bebased on the applicable API standard for centrifugalcompressors (API Standard 617) and reciprocatingcompressors (API Standard 618) and must be mutually agreedupon by the vendor, the buyer, and the Saudi Aramco Engineer.

A mechanical running test is an operational test of the

compressor that is conducted at the vendors facilities. Themechanical running test must be of four hours in duration forboth centrifugal and reciprocating compressors. The four-hourmechanical running test allows compressor components, suchas bearings and rotors, to become thermally stable. SAES-K-403 requires that a mechanical running test must be performedon all reciprocating compressors and that this test must bewitnessed for reciprocating process gas compressors. SAES-K-


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402 requires that a mechanical running test must be performedand witnessed on all centrifugal compressors. When a sparerotor is purchased, the rotor must be installed and run in aseparate test prior to the job rotor test.

For reciprocating compressors, the mechanical running testproves the mechanical operation of all of the auxiliaryequipment as well as the compressor, reduction gears (ifapplicable), and the driver. The compressor does not have tobe pressure-loaded for this test. The test is not acceptable ifany repair or replacement is required to correct mechanical orperformance deficiencies that are identified during themechanical running test. The test must be rerun after therepairs or corrections are completed.

For centrifugal compressors, the following requirements must besatisfied prior to the performance of the mechanical runningtest:

The shaft seals and bearings that were specified with thecompressor must be installed and used in the machine forthe mechanical running test.

The oil pressures, the oil viscosities, and the oiltemperatures must be at the same operating values as theoperating values that are recommended in themanufacturers operating instructions for the specific unit

under test. The oil filtration must be ten microns nominal orbetter.

All joints and connections must be checked for tightness.Any leaks must be corrected prior to the mechanical runningtest.

Facilities must be installed to prevent the entrance of oil intothe compressor during the test. These facilities must be inoperation throughout the test.

All warning, protection, and control devices must becalibrated to the their alarm, shutdown, or relief set points.

Any auxiliary gear units that are supplied with thecompressor must be included in the mechanical running test.The mechanical test should include the coupling that is to beinstalled on the compressor. If the inclusion of the jobcoupling is not practical, the mechanical running test mustbe performed with coupling-hub moment simulators in place.When all of the tests are complete, the moment simulators


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must be furnished as part of the special tools for thecompressor.

The radial vibration and axial position transducer sensors,

signal conditioners, and connecting cables that are to besupplied with the compressor must be used in the test. If thevendor does not furnish the vibration monitoring equipmentor if the equipment is not compatible with the test shopreadout equipment, shop equipment and readouts that meetthe accuracy and calibration requirements of the applicable

API Standard must be used.

The compressor should be started and operated at speedincrements of approximately 10% from zero to the maximumcontinuous speed. The compressor is run at the maximum

continuous speed until the bearing and lube oil temperaturesand the shaft vibrations have stabilized. Once the bearing andlube oil temperatures and the shaft vibrations have stabilized,the speed is increased to the trip speed, and the compressor isoperated for a minimum of 15 minutes. After 15 minutes, thespeed of the compressor is adjusted to the maximumcontinuous speed, and the equipment is run for a minimumduration of the test (four hours) at the maximum continuousspeed. The mechanical operation of all equipment being testedand the operation of the test instrumentation must besatisfactory during the test. During the four-hour test, radial

shaft vibration, bearing pad temperature, lubrication supply, andreturn temperatures and flow must be measured. The innerseal-oil leakage rate must be measured at each seal. The lubeoil and seal oil inlet pressures and temperatures should bevaried through the range that is permitted by the compressorsoperating manual.

Processed from unfiltered transducer output signals,measurements of radial shaft vibration and axial position mustbe recorded, and they must not exceed the applicable vibrationlimits throughout the test. While the mechanical test is beingconducted, vibration sweep readings must be recorded forvibration amplitudes at frequencies other than synchronous. Asa minimum, these sweep readings must cover a frequencyrange from 0.25 to 8 times the maximum continuous speed, butthey must not exceed 90,000 cycles per minute (1500 Hertz).Polar plots that show the synchronous vibration amplitude (interms of radial shaft vibration), phase angle, and phase shiftversus rotational speed must be made before and after the four-


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hour test. The speed range that is covered by the plots must befrom zero to the specified driver trip speed. The critical speedsof the compressor must be verified during the mechanicalrunning test. Taped recordings of all real-time vibration data

should be made during the mechanical running test. Theserecordings provide the initial data for vibration analysis.

An inspection that includes the dismantling, the inspection, andthe reassembly of the compressor, the gear, and the driver mustbe made after satisfactory completion of the mechanical runningtest. A bearing inspection must be completed. All bearingsmust be removed, inspected, and reassembled after completionof the mechanical running test. Shaft seals should be removedfor inspection. If minor scuffs or scratches occur on bearings oron shaft seal surfaces, minor cosmetic repairs are not a cause

for rerunning the test.

The rotor dynamics of a compressor include the followingdifferent areas and considerations:

The performance of a lateral analysis.

The performance of a torsional analysis.

The performance of assembly vibration testing andbalancing.

When an exciting frequency is applied to a rotor-bearing supportsystem that corresponds to the natural frequency of the rotor-bearing support system, the system may be in a state ofresonance. A resonating rotor-bearing support system will haveits normal vibration displacement amplified.

The magnitude of amplification and the rate of phase shift(phase-angle change) are related to the amount of damping inthe rotor-bearing support system and the mode shape that istaken by the rotor as it deflects. The mode shapes for deflectionare commonly referred to as the first rigid (translatory orbouncing) mode, the second rigid (conical or rocking) mode, thefirst bending mode, the second bending mode, and the thirdbending mode. An exciting frequency may be less than, equalto, or greater than the rotational speed of the rotor. Thefollowing are some of the sources of exciting frequencies thatmust be considered:

Unbalance in the rotor system.


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Oil-film instabilities (whirl).

Internal rubs.

Blade, vane, nozzle, and diffuser passing frequencies.

Gear-tooth meshing and side bands.

Coupling misalignment.

Loose rotor-system components.

Friction whirl.

Boundary-layer flow separation.

Acoustic and aerodynamic cross-coupling forces.

Asynchronous whirl.

The magnitude of the amplification is called the rotoramplification factor. The rotor amplification factor (AF) isdetermined through use of the following formula and the bodeplot that is shown in Figure 1:

12

1c

NN

NAF

=

The bode plot is a graph of amplitude versus the rotor speed (inrevolutions per minute) and phase (between the shaft reference

mark and peak vibration) versus rotor speed (in revolutions perminute. The polar plot is a graph of amplitude versus phase fora range of compressor speeds Figure 1 represents an actualcentrifugal compressor rotor response. The specific points ofinterest on the bode and polar plots are identified. A rotorresponse plot provides the following information:

The rotors first critical speed in revolutions per minute (Nc1).

The rotors initial (or lesser) speed (N1). The initial speedoccurs at the first peak-to-peak amplitude that is equal to0.707 times the peak-to-peak amplitude at the critical speed

(Ac1). The rotors final or greater rotational speed (N2) occurs after

the displacement at the first critical speed. The value ofpeak-to-peak displacement at N2is equal to 0.707 of peak-to-peak displacement at N1.

The peak-to-peak amplitude (Ac1) at the rotors first criticalspeed (Nc1).


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Figure 1. Typical Rotor Response Plot


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If the rotor amplification factor is greater than or equal to 2.5, thevibration frequency at which resonance occurs is called critical.The rotational speed at which the resonance occurred is calleda critical speed. A critically damped system is a system that has

an amplification factor of less than 2.5.

Critical speeds for compressors must be determined analyticallythrough use of a damped, unbalanced, rotor response analysis,which must be confirmed by test-stand data. Resonances thatoccur within the specified operating speed range of thecompressor must be critically damped. Any operating speedthat should be avoided as a critical speed must be included inthe operating and maintenance instructions for the compressor.The critical speeds of the driver must be compatible with thecritical speeds of the compressor, and the combination must be

suitable for the operating speed range.

It is the vendors responsibility to provide a damped,unbalanced-response, lateral analysis for the compressor inorder to ensure acceptable amplitudes of vibration at any speedfrom zero to trip. The effects of other equipment in the trainshould be included in the damped, unbalanced-responseanalysis. The following considerations should be included in thedamped, unbalanced-response analysis:

Support stiffness (base, frame, and bearing housing), mass,and damping characteristics. These characteristics mustinclude the effects of rotational speed variations.

Bearing lubricant-film stiffness and any damping changesthat are due to speed, load, preload, oil temperatures,accumulated assembly tolerances, and maximum tominimum bearing clearances.

Rotational speeds (starting speeds, operating speed andload ranges, trip speed, and coast-down speeds). (Anyspecial speeds, such as test condition speeds, should alsobe included.)

Rotor masses, which include the mass moment, thestiffness, and the damping effects of the coupling halves.(Examples of the damping effects are accumulated fittolerances, fluid stiffening and damping, and frame andcasing effects.)

Asymmetrical loading. (Examples of asymmetrical loadingare partial arc admission, gear forces, side streams, andeccentric clearances.)


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As a minimum, the damped, unbalanced-response analysismust include the following:

A plot and identification of the mode shape at each resonant

speed (critically damped or not) from zero to trip. The nextmode that occurs above the trip speed must also beincluded in the plot.

The frequency, phase, and response amplitude data thatwere based on measurements processed from the vibrationprobe locations over the range of each critical speed.

For each response, diagrams that indicate the phase and themajor-axis amplitude at each coupling engagement plane,the centerlines of the bearings, the locations of the vibrationprobes, and each seal area throughout the machine. The

minimum design diametral running clearance of the sealsmust also be indicated.

An additional plot of the unbalance and location (usually thecoupling) that will be used for shop testing. This additional,unbalanced-response plot must include the effects of anytest stand conditions or test seals that may be used toperform the shop verification test.

A stiffness map of the undamped rotor response from whichthe damped unbalanced response analysis was derived.This plot should show frequency versus support systemstiffness. The calculated support system stiffness curvesare superimposed.

The damped unbalanced response analysis must indicate thatthe compressor, in the unbalanced condition, will meet thefollowing acceptance criteria:

If the amplification factor is less than 2.5, the response isconsidered critically damped, and no separation margin isrequired.

If the amplification factor is between 2.5 and 3.55, aseparation margin of 15% above the maximum continuousspeed and 5% below the minimum operating speed isrequired.

If the amplification factor is greater than 3.55 and if thecritical response peak is below the minimum operatingspeed, the required separation margin as a percentage of


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minimum speed is determined by the following equation:

+=3AF

684100SM

Where:

SM = Separation Margin

AF = Amplification Factor

If the amplification factor is greater than 3.55 and if thecritical response peak is above the trip speed, the requiredseparation margin, as a percentage of maximum continuousspeed, is determined by the following equation:

1003AF

6126SM

=

Where:

SM = Separation Margin

AF = Amplification Factor

A shop verification of the unbalanced-response analysis mustbe performed. The actual responses are the criteria used to

confirm the validity of the damped unbalanced responseanalysis. The shop verification is performed on a test stand witha rotor unbalanced magnitude of at least two times and no morethan eight times the specific unbalanced limit, typically placed atthe coupling. The actual critical speed responses are recordedon the test stand. The dynamic response of the machine on thetest stand is a function of the test conditions. The test resultsshould be obtained at the conditions of pressure, temperature,speed, and load that are the expected in the field; otherwise, thetest stand results may not be comparable with what occursduring actual operation in the field.

The performance of a torsional analysis includes adetermination of the excitations of torsional resonances of thecompressor. Excitations of torsional resonances should beconsidered in the dynamics analysis. These excitations may beproduced from any of the following partial list of sources:

Gear problems, such as unbalanced gears and pitch linerunout.


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Gas pressure forces or unbalanced mass in connecting rodof reciprocating compressors.

Start-up conditions that include speed detents that are under

the inertial impedances as well as other torsionaloscillations.

Torsional transient, such as startups of synchronous and/orvariable frequency electric motors.

Any greater torsional resonances, including the naturalfrequencies, that are a product of the complete train must be atleast 10% above or 10% below any possible excitationfrequency that exists within the speed range of minimum tomaximum continuous speed. Torsional resonances are calledtorsional criticals if they occur at frequencies that are twice thecompressors running speeds or greater, and they should beavoided. If the compressors torsional resonances arecalculated to be a multiple of the running speed and if all effortsto remove the critical from within the limiting frequency rangehave been exhausted, a stress analysis must be performed todemonstrate that the resonances have no adverse effect on thecomplete compressor train.

The major components of the rotating element of a compressor(the shaft, balancing drum, and impellers) must be vibration-tested and dynamically-balanced. When a bare shaft with a

single keyway is dynamically-balanced, the keyway must befilled with a fully crowned half-key for an initial balance. Thisinitial balance correction to the shaft must be recorded.

The rotating element (rotor) must be multi-plane, dynamicallybalanced during the assembly of the compressor. Two of themajor components that make up the rotating element may beadded to the rotating element prior to completion of the dynamicbalancing. Any corrections that must be made to the rotatingelement to correct an unbalance condition must be applied tothe components that were added to the rotating element. After

the compressor is completely assembled, minor corrections ofother components that were added to the assembly may berequired. These minor corrections will be determined during thefinal trim balancing of the completely assembled element.

Residual unbalance is the amount of unbalance that remains ina rotor after the rotor has been balanced. For dynamiccompressors, API Standard 617, Appendix D provides the


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specific procedures and calculations for determining the residualunbalance of a dynamic compressor. The following equation isused to calculate the maximum allowable residual unbalanceper plane for a compressor:

4W/NUmax =

Where:

Umax = Amount of residual unbalance, in ounce-inches(gram-millimeters).

W = The journal static weight load, in pounds(kilograms).

N = The maximum continuous speed, in revolutionsper minute.

After the balancing machine readings indicate that the rotor hasbeen balanced to within the specified tolerances, a residualunbalance check should be performed before the rotor isremoved from the machine. To perform a residual unbalancecheck (multiplane balancing), a known trial weight is attached toone of the balance planes of the rotor, and a balance check isperformed. The weight is moved around the rotor in six ortwelve equal increments and a balance check is performed.The trial weight is moved to the next balance plane, and the testis repeated until all of the balance planes have been tested.The balance check readings are plotted on a polar plot, and the

amount of residual unbalance is calculated. If the specifiedmaximum allowable residual unbalance has been exceeded inany balance plane, the rotor must be balanced more precisely,and the residual-unbalance check must be repeated.

The peak-to-peak amplitude of unfiltered vibration in anyspecific plane is tested during the testing of the balanced rotor.With a balanced rotor operating at its maximum continuousspeed, the peak-to-peak amplitude of unfiltered vibration that ismeasured on the shaft adjacent and relative to each radialbearing must not exceed its calculated limitation or 2.0 mils (50

micrometers) on any plane, whichever is less. The peak-to-peakamplitude of unfiltered vibration limitation is calculated throughuse of the following formula (for U.S. customary units):

N

12,000A=

Where:


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A = The amplitude of unfiltered vibration, in mils(micrometers) peak-to-peak.

N = The maximum continuous speed, in revolutions

per minute.For any speed that is greater than the maximum continuousspeed, the vibration limit is a comparison to the maximumvibration value that is recorded at the maximum continuousspeed. The vibration for any speed that is greater than themaximum continuous speed must not exceed 150% of thevibration value that is recorded at the maximum continuousspeed.

If the vendor can demonstrate that electrical runout ormechanical runout is present in the rotor system, a maximum of

25% of the peak-to-peak amplitude of unfiltered vibration thatwas calculated from the above formula or 0.25 mil (6.4micrometers), whichever is greater, may be subtracted from thevibration signal that is measured during the factory testing. Theelectrical and mechanical runout are determined by rotation ofthe rotor in V-blocks at the journal centerline while measuringthe runout. The runout measurement is measured with anoncontact proximity probe (for electrical runout) and with a dialindicator (for mechanical runout). The runout measurement istaken for the full 360 degrees of rotation. The noncontactproximity probe is located at the normal probe location, and the

dial indicator is located one probe tip diameter on either side ofthe noncontact proximity probe. The electrical runout andmechanical runout readings are recorded. The electrical runoutand mechanical runout readings must be supplied by the vendorin the mechanical test report.

Gas Leakage Test

After the mechanical running test is completed, each completely

assembled, centrifugal compressor casing that is intended fortoxic or flammable gas service must have a gas leakage test asspecified in API 617. The gas leakage test must be witnessed.The requirements of API 617 may require two separate tests asdescribed in the following text to accomplish the gas leakagetest.

The casing (including the end seals) is pressurized with an inert


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gas to the maximum sealing pressure or the maximum sealdesign pressure. The test is considered satisfactory when nocasing or casing-joint leaks are observed or detected.

When specified, the casing (with or without the end sealsinstalled) is pressurized to the rated discharge pressure and isheld at this pressure for a minimum of 30 minutes. After 30minutes, a soap-bubble test (or another approved test) isperformed to check for gas leaks. The test is consideredsatisfactory when no casing or casing-joint leaks are observedor detected.

Performance Test

In accordance with SAES-K-402 for centrifugal compressors, asa minimum, a performance test must be specified andwitnessed for each centrifugal compressor duty. For a series ofidentical units, only one unit needs to be performance tested.Tests must be in accordance with ASME Power Test Code 10-1965 (compressors and exhausters), Class I, II, or III. Testsmust be to Class III specifications unless otherwise specified.Class I or Class II tests must be considered for medium to highdischarge pressures (500 psia) where rotor instability that is dueto high gas densities could be encountered or where

compressors are located on an offshore platform. In thesecircumstances, Saudi Aramcos Engineer must be consultedconcerning advisability of Class I or II full load, full pressuretests. The extra costs of such tests, as compared to a Class IIItest, must be weighed against the cost (and delay) to correctany malperformance after the compressors are installed.

ASME Power Test Code (PTC10-1965) has defined thefollowing three classes of performance tests:

Class I, which is a test run on the design gas at near designconditions. This test generally applies to air compressors.

Class II, which covers tests when using the design gas is notpractical. Both test and design gas must closely followperfect gas laws.

Class III, which is similar to test 2 in that a different gas isused for the test; however, in this test, the gas does notfollow the perfect gas law.


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When a performance test is conducted through use of asubstitute gas, the test must be performed at an equivalentspeed. In accordance with ASME PTC-10, when operating atthe equivalent speed, the test parameters must agree with the

corresponding field parameters. ASME PTC-10 includes therequired tables, the procedures, and the calculations that arenecessary to determine the equivalent speed and to correct thetest results to actual field conditions.

A minimum of five points that include surge and overload mustbe taken at normal speed. For variable-speed machines,additional points may be specified. Head and capacity shouldhave zero negative tolerance at the normal operating point (orother points as specified). The horsepower at this point shouldnot exceed 104% of the specified value. The compressor test

must show that the compressor is suitable for continuousoperation at any capacity at least 10% greater than thepredicted approximate surge capacity that is designated on thedata sheets.

For constant-speed compressors, the head should be within therange of 100% to 105% of the normal head. The horsepowerwill be based on the required normal head and capacity.

Unless otherwise specified, the performance test should beconducted through use of only one contract rotor.

Field test procedures should be in accordance with ASMEPTC10-1965, Compressors and Exhausters, within practicallimits. Tests should not be conducted until it is certain that thecompressor has reached equilibrium, with all parameters asclose as possible to those parameters that are anticipated inactual service.

All pressure and temperature instrumentation must be properlycalibrated. The ASME code provides guidelines forinstrumentation of the external flanges of the compressor and

the flow measuring sections. This instrumentation will provideadequate readings at the compressor flanges and flow-measuring devices. Temperatures should be measured throughuse of a thermocouple or an RTD system. The sensitivity and

readability of the temperature measuring device should be .5Fand should have an accuracy within 1F. Pressure readingsduring testing should be at approximately mid scale. Thepressure gage sensitivity should be about .25% with a .5%


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maximum error of full scale when reading pressures that aregreater than 20 psig. When pressures are less than 20 psig, avertical manometer should be used unless disallowed for safetyreasons.

A gas sample should be taken at the top of the suction and thedischarge of the compressor at the beginning and the end of thetest. To avoid condensation in the sample, the gas samplemust be analyzed at a temperature that is equal to or greaterthan the expected field conditions. The gas sample should beanalyzed by a gas chromatograph. The equipment speedshould be determined through use of two independent phasereference transducers. Mass flow rates are measured throughuse of the process flow indicator but should be verified throughcalculations; therefore, metering device upstream temperature,

upstream pressure, and differential pressure must also berecorded.

If field tests are conducted to confirm that the guaranteedconditions on new equipment are met, the acceptancetolerances are the same as noted for the performance tests.When field tests are conducted on existing equipment todetermine whether inspection is required, a reduction ofpolytropic head and/or efficiency of 10% or greater from theperformance test results (at rated flow) is a sound basis forrecommending internal compressor inspection.

String Test

As specified in SAES-K-402, a string test is used for longequipment trains, for off-shore installations, and in situationswhere early detection of equipment malfunction is necessary.The following system components are tested as a unit:

Driver

Gear

Compressor(s)

Oil Systems (Lube Oil and Seal Oil Systems)

String tests require prior agreement by Saudi AramcosEngineer, and they must be witnessed. String tests areperformed in addition to separate tests of individualcomponents. Torsional vibration measurements are to beperformed to verify data sheets.


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Post-Test Inspection

In accordance with SAES-K-402 for centrifugal compressors, all

bearings and seals (except labyrinth types) must be removedand inspected after the completion of the mechanical runningtest. Additional dismantling, inspection, and re-assembly of thecompressor should be considered an optional extra forapplication only in special circumstances. The merits of a post-test inspection of the casing internal should be evaluatedagainst the benefits of shipping a unit with proven mechanicalassembly and casing joint integrity.

In accordance with SAES-K-403 for reciprocating compressors,dismantling of the compressor after the mechanical running test

should be requested only on a compressor that is not of aproven design. This dismantling and inspection is other thanany dismantling and inspection that is required by evidence of amalfunction during the mechanical running test.


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DETERMINING DYNAMIC COMPRESSOR ACCEPTABILITY

When determining a dynamic compressors acceptability,several characteristics of the compressor must be considered.The relationships between the inlet volume flow, head, speed,efficiency, and power of a dynamic compressor are oftenreferred to as the compressors characteristics. The actualcompressor characteristics are compared to the compressorsvendor-guaranteed characteristics. The compressor mustperform within the specified tolerances of the guaranteedcharacteristics.

Figure 2 is a typical multi-stage centrifugal compressorcharacteristic curve. The curve is a plot of the inlet volume inpercent versus the head in percent and the horsepower in

percent. Speed and efficiency lines are plotted on the curve toprovide the compressors characteristics. The speed that isshown on the curve ranges from 70% to 110%. The efficiencythat is shown on the curve ranges from 83% to 100% of thepeak efficiency, with the approximate surge line drawn in at alow volume and efficiency.

As an example, if a compressor has a rated efficiency of 75%,the efficiency will be 75% of the 100% line and 62.25% on the83% line.


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Figure 2. Typical, Multi-Stage, Centrifugal Compressor Characteristic Curve


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Figure 3 is a typical axial compressor characteristic curve. Thecurve is a plot of the design volume in percent versus the designcompression ratio in percent. Speed and efficiency lines areplotted on the curve to provide the compressors characteristics.

The speed that is shown on the curve ranges from 75% to105%. The efficiency that is shown on the curve ranges from80% to a maximum efficiency of 100% of peak efficiency.

Figure 3. Typical Axial Compressor Characteristic Curve

Once the decision has been made as to which type ofcompressor is to be used in a given application, eachcompressor is tested against its vendor-guaranteedcharacteristics to determine compressor acceptability. Thisremainder of this section of the Module will examine thefollowing areas that Saudi Aramco Engineers must considerwhen determining the acceptance of dynamic compressors:

Acceptability Criteria (31-SAMSS-001)

Calculating Inlet Flow Volume

Polytropic Calculations

Calculating Pressure Ratio from Head

Calculating Horsepower and Efficiency

Use of Fan Laws to Find Operating Point at Different Speeds


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Acceptability Criteria (31-SAMSS-001)

The following account is cited from 31-SAMSS-001 (forcentrifugal compressors), which adopts and specifiesexceptions to API Standard 617. Centrifugal compressorsystems must be supplied by vendors who are qualified byexperience in manufacturing the proposed units. To qualify, thevendor must have manufactured, at the proposed location ofmanufacture, at least two compressors of comparable speed,power rating, and discharge pressure for a gas of comparablecharacteristics. These compressors must have been inoperation for at least one year, and they must be performingsatisfactorily.

Calculating Inlet Flow

The performance curves that are supplied by the manufacturerand the machines performance are usually based on the actualvolume flow at the suction of the compressor. The calculationsto determine these performance curves have been discussed inthe Volumetric Flow and Mollier Method sections of Module212.02. It is important that the Mechanical Engineer understandthat the process data are usually given in SCFM or lb./hr andthat the process data must be converted to ACFM in order to

determine compressor performance.

For the purpose of performance calculations, compressorcapacity is expressed as the actual volumetric quantity of a gasat the inlet to each stage of compression on a per minute basis(ICFM). All centrifugal compressors are based on actual flow,which is converted to inlet or actual cubic feet per minute. Thisconversion is done because a dynamic compressors (axial andcentrifugal) produced head is a function of inlet gas velocity.The inlet velocity is derived by the division of volume flow byblade area. The following equation is used to determine inlet

flow (Q1) in actual or inlet cubic feet per minute (ICFM) if theinlet flow is known in standard cubic feet per minute (SCFM):

11

1

ZR520

T

P

14.7SCFMICFM

=

Where:

P1 = Inlet pressure (psia)


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T1 = Inlet temperature (R) (R = F = 460)

Z1 = Inlet compressibility factor

The following example will determine the actual ICFM of acompressor that has suction of 60,000 SCFM at an inlet

pressure (P1) of 100 psia, an inlet temperature (T1) of 100F,and an inlet compressibility factor (Z1) of 1.0.

/minft9525

1.01.08.14760,000

1.0R520

R560

psia100

14.760,000

ZR520

T

P

14.7SCFMICFM

3

11

1

=

=

=

=

The following equation is used to determine inlet flow (Q1) inactual or inlet cubic feet per minute (ICFM) if the inlet flow isknown in weight flow (mass flow) in lb/min:

W

ICFM=

Where:

w = Mass flow (lb./min)

= Density (lb./ft3)

The following equation is used to determine density ():

11

1

ZxT

P

x28.95

MW

x2.7=

Where:

MW = Molecular weight

P1 = Inlet pressure (psia)


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T1 = Inlet temperature (R) (R = F = 460)

Z1 = Inlet compressibility factor

The following example will determine the actual ICFM ofa compressor that has weight flow () of 3600 lb./min atan inlet pressure (P1) of 100 psia, an inlet temperature

(T1) of 100F, a molecular weight (MW) of five, and acompressibility factor (Z1) of 0.98.

/minft42,353

lb/ft0.085

lb/min3600

pQ

0.085

0.1820.1722.7

0.98560

100

28.95

52.7

ZT

P

28.95

MW2.7

3

3

11

1

=

=

=

=

=

=

=

Polytropic Calculations

The actual compression path does not follow any reversibleprocess (isothermal, isentropic, or polytropic). The actual

compression path is most closely approximated by thepolytropic process in which PVn = a constant. In such cases,polytropic calculations must be used. The polytropic head isobtained through use of the following equation:

=

1P

P

1)/n(nMW

TRZHeadPolytropic

1)/n(n

1

21univavg


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Where:

T1 = Inlet temperature

P1 = Inlet pressure

P2 = Discharge pressure

n = Polytropic exponent


Zavg = Average compressibility factor

Runiv = Universal gas constant (1545.32 ft-lbf/lbm-Mol-

R)

The polytropic exponent () factor may be found from theequation:

P

1

k

1k

n

1n

=

Where:

k = Isentropic exponent

P = Polytropic efficiency

Also, the relationship between discharge pressure andtemperature for an ideal gas polytropic process is useful duringfield testing when polytropic efficiency is not known. Thisrelationship can be stated as the following equation:

1)/n(n

1

2

1

2

P

P

T

T

=

Where:

T2 = Discharge temperature

This equation can be rewritten as follows:


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=

1

2

1

2

P

P

Ln

T

TLn

n

1n

For real gases, n can be found from the following relationship:

=

1

2

1

2

ln

ln

n

Where:

1 = Inlet density

2 = Outlet density

Calculating Pressure Ratio from Head

The primary variable in calculating head required is pressure, P2and P1. The plot of pressure ratio versus flow rate will be similarto head versus flow rate. Pressure ratio is calculated throughuse of the following equation:

1

2P

P

Pr =

Where:

rP = Pressure ratio

P1 = Inlet pressure


The pressure ratio can be calculated from the head through useof the following equation:

1n

n

1univavg

p

P 1n

1n

TRZ

MWxHr

+

=


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Where:

Hp = Polytropic head



Runiv = Universal gas constant


n = Polytropic exponentFor the same compressor that is operating at the same flow and

speed, rPwill change if MW, Z, -1/, or /-1 changes.

Calculating Horsepower and Efficiency

The efficiency of a thermodynamic system is stated as the ratioof the work output of the system (head) to the work input to thesystem (shaft power).

The difference between head and work is the amount of lossesthat are internal to the machine due to such conditions asfriction and windage. These losses show up as heat, and theyadd to the discharge temperature.

These losses include losses that are external and internal to the

main flow path. Losses that are external to the main flow pathinclude losses such as windage losses, disk friction losses, andleakage losses. Losses that are internal to the main flowpathare actual losses of blade input energy, and they include thefollowing:

Skin friction

Blade loading and diffusion

Incidence angle

Exit mixing losses

Clearance losses

Horsepower is the rate of doing work. If a compressor is lifting aweight of gas to a given head (H) at a specific rate (M),horsepower would be calculated as follows:


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33,000x

MxHGHP=

Where:GHP = Gas horsepower

H = Head (ft-lbf/lbm)

M = Weight flow (lb./min)

= Efficiency

If polytropic head (Hpoly) is used in the equation, polytropic

efficiency (poly) must also be used. If isentropic head (Hisen) isused in the equation, isentropic efficiency (isen) must also be

used.

Gas horsepower is not the true input horsepower to thecompressor. Mechanical and hydraulic losses must beconsidered in order to determine the true input horsepower orbrake horsepower (bhp). Typical losses to be considered are asfollows:

Bearing losses

Seal losses

Friction losses

Other losses, such as radiation losses, labyrinth seal losses,and recirculation due to balancing devices, may typically beignored in calculating bhp. Bhp can be calculated as follows:

bhp = GHP + mechanical losses

The mechanical losses are typically noted on the compressordata sheets. If the mechanical losses are not available, anestimate of 20 hp may be used for seal losses, and an estimateof 50 hp may be used for bearing horsepower. The estimatedhorsepower values may vary greatly due to bearing load, speed,and oil temperature.

As the following equation shows, adiabatic or isentropicefficiency uses isentropic relationships to define head (usefulwork) and total work input.


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( )

12

1)/k(k

121

adTT

1/PPT

=

Where:

ad = Adiabatic efficiency


T2 = Discharge temperature

P1 = Inlet pressure


k = Isentropic exponent

The overall adiabatic efficiency is useful as a measure of theoverall performance of a compressor in the determination ofpower; however, adiabatic efficiency is not always a trueindication of efficiency in reference to internal losses. Figure 4illustrates this point. Because isentropic work is proportional totemperature rise (Wad= cpDT), the distance from point 1 to point2adis proportional to the adiabatic work that is required tocompress the gas from P1to P2. The actual work, however, isproportional to the vertical distance from point 1 to point 2.


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Figure 4. Adiabatic Versus Polytropic Process

The polytropic equation represents the true aerodynamicefficiency of a compressor for compression of an ideal gas.There are, however, limitations to this equation. Real gases donot always have a constant k value. The value of k for somegases at discharge conditions can vary significantly from the kvalue at suction conditions. The enthalpy (or Mollier) equationis the most accurate method to calculate the aerodynamicefficiency for any condition. In some cases, the enthalpy (orMollier) is the only equation that can provide accurate results. If

the average value of k is known and, if is determined bycalculation, pcan be determined by the following:


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(BTU/lbm)h(BTU/lbm)h

lbf/BTU)(ft778

lbf/lbm)(ftH

InputWork

Head

1)/n(n

1)/k(k

12

p

avgavg

P

=

=

=

Where:

P = Polytropic efficiency

Hp = Polytropic head (BTU/lb.)

h2 = Discharge enthalpy (BTU/lb.)

h1 = Inlet enthalpy (BTU/lb.)

kavg = Average isentropic coefficient

n = Polytropic coefficient

Use of Fan Laws to Find the Operating Point at Difference Tip Speeds

The fan laws for centrifugal compressors are similar to theaffinity laws for centrifugal pumps. The following equationsshow the relationship between the volume flow rate (Q), thehead (H), the horsepower (bhp), and compressor speed (N):

Equation 1

=

1

212

N

NQQ

Equation 22

1

212

N

NHH

=

Equation 3


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3

1

212

N

Nbhpbhp

=

Where:

Q = Suction flow, actual

H = Polytropic head

bhp = Brake horsepower

N = Speed, rpm

As indicated in Equation 1, the performance of a centrifugalcompressor at speeds other than design speed is such that thecapacity or flow rate will vary directly as the speed varies. Asindicated in Equation 2, the head that is developed will vary as

the square of the speed varies. As indicated in Equation 3, thehorsepower will vary as the cube of the speed varies. As thespeed of the compressor deviates from the design speed, theerror of these laws increases. The fan laws only accuratelyapply to single-stages with very low compression ratios.

These laws can be used to estimate the performance at onespeed if the performance at another speed is already known.The accuracy of the fan laws decreases with increasingcompressor stages and gas density; therefore, the actualperformance prediction at off-design speeds must be obtained

from the compressors vendor.

If the curve at speed N1is known, these relationships are usedto draw the head and horsepower curves at speed N2, as shownin Figure 5. Starting with any point on the head curve at speedN1(point A1), both the head (H2) and the flow (Q2) are calculatedby equations 1 and 2. This calculation gives an equivalentoperating point on the curve for speed N2(point A2). A series ofthese points defines the curve for N2. Similarly, for thehorsepower curve that is shown in Figure 6, the horsepower(bhp2) and the flow (Q2) are calculated to obtain the equivalent

operating points.


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Figure 5. Head Curve

Figure 6. Horsepower Curve


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DETERMINING POSITIVE-DISPLACEMENT COMPRESSORACCEPTABILITY

This section of the Module will examine the following areas thatSaudi Aramco Engineers must consider when determining theacceptance of positive-displacement compressors:

Acceptability Criteria

Calculating Capacity

Calculating Discharge Temperature

Calculating Power

Acceptability Criteria (31-SAMSS-002/31-SAMSS-003)

The following account is in accordance with 31-SAMSS-002 (forpackaged reciprocating plant and instrument air compressors),which adopts and specifies exceptions to API Standard 680.Reciprocating compressors must be supplied by vendors whoare qualified in manufacturing the proposed units. To qualify,the vendor must have manufactured, at the proposed point ofmanufacture, at least two compressors of identical frame size.These compressors must have been in service in desertenvironment conditions (as specified in Section 2.1.13 of 31-

SAMSS-002 and in SAES-A-112) for at least one year, and theymust be performing satisfactorily.

In addition to the design criteria of 20 years of service life,compressors must be suitable for a minimum period of 10,000hours of uninterrupted operation between planned maintenanceshutdowns.

The vendor must advise Saudi Aramcos Engineers of the flowrate, the outlet temperature, and the inlet pressure that arerequired at design conditions for Saudi Aramcos specifiedcoolant inlet temperature (specified on the data sheet).

The proposal and the operating instructions must specifymaximum and minimum operating conditions of the unit aslimited by pressure, temperature, and other conditions thatcould shorten the life of the machine. The vendor must furnishany required protective devices that are used to preventdamage to the equipment.

In accordance with 31-SAMSS-003, reciprocating compressors


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for process air or gas service must be supplied by vendors whoare qualified in manufacturing the proposed units (31-SAMSS-003 adopts and specifies exceptions to API Standard 618). Toqualify, the vendor must have manufactured, at the proposed

point of manufacture, at least two compressors of identicalframe size, speed, power rating, and discharge pressure for agas of comparable characteristics. These compressors musthave been proven in service in a desert environment for at leastone year, and they must be performing satisfactorily.

Calculating Capacity

The value for the capacity of a positive-displacementcompressor is used to determine other pertinent compressoroperating characteristics. The theoretical capacity of thepositive-displacement compressor is used to compare againstthe actual measured capacity of a compressor that is installed ina system. A large difference between the calculated theoreticalcompressor capacity and the measured compressor capacityindicates that there may be compressor component degrada-tion. The following equation is used to calculate the theoreticalmaximum capacity of a reciprocating compressor cylinder:

VExDISPxZ

Zx

T

Px0.0509Q

s

std

s

s=

Where:

Q = Theoretical maximum capacity in millionstandard cu ft per day (mmscfd) at 14.7 psia

and 520R

PS = Suction pressure in psia

TS = Suction temperature in R

Zstd = Compressibility factor at standard conditions

ZS = Compressibility factor at suction conditions

DISP = Cylinder displacement in cu ft per minute(cfm)

VE = Volumetric efficiency

To determine the capacity of a single stage that has more thanone cylinder, the capacity value should be multiplied by thenumber of cylinders in the stage.


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VolumetricEfficiency

The critical portion of the theoretical capacity equation is thetheoretical volumetric efficiency, as defined as:

VE = 1 - C(R1/k- 1)

Where:

C = Percent clearance as a decimal fraction ofdisplaced volume

R = Pressure ratio across the cylinder (dischargepressure divided by suction pressure, psia)

k = Isentropic volume exponent at operatingconditions (the specific heat ratio for ideal gas,Cp/Cv)

An alternate equation for determining the theoretical volumetricefficiency is:

VE = 100 - R- C

1R

Z

Z1/k

d

s

Where:

Pd = Discharge pressure in psiaPs = Suction pressure in psia

Zd = Compressibility factor at discharge conditions

Zs = Compressibility factor at suction conditions



C = Cylinder clearance as a percentage

Volumetric efficiency is the suction volume flow rate divided bythe displacement. For a reciprocating compressor, thetheoretical volumetric efficiency is considerably less than 100%because of clearance volume, valve losses, piston ring leakage,and packing losses. Clearance volume is that portion of thecylinder that is not swept by the piston. At the end of a


42/62




discharge stroke, the clearance volume is filled with a gas atdischarge pressure. During the subsequent suction stroke, thisgas begins to expand. The suction valve does not open untilthe gas in the clearance volume expands from the discharge

pressure to the suction pressure. After this expansion, the gasis admitted to the cylinder until the end of the suction stroke;however, the gas is admitted during only 70% to 80% of thetotal suction stroke. The amount of lost suction volumedepends on the compression ratio, the properties of the gas,and the amount of clearance volume.The volumetric efficiency of an operating compressor iscalculated to determine whether the valves, pistons, andpacking are properly operating. An actual volumetric efficiencythat is significantly less than the theoretical value indicates thatthe valves, the piston rings, and/or the packing are leaking and

that maintenance is required. The actual volumetric efficiencycan be determined by the following equation:

VE = 1 - L - C (R1/k

- 1)

Where:

C = Percent clearance as a decimal fraction ofdisplaced volume



L = Loss correction factor as a decimal fraction

or

VE = 100 - R- L - C

1R

Z

Z 1/k

d

s

Where:C = Cylinder clearance as a percentage

L = Loss correction factor as a percentage

The loss correction factor (L), which accounts for the valvepacking losses and the piston ring losses, can be obtained fromFigure 7. The loss correction factor is determined by locating


43/62




the compression ratio value on the x axis. The compressionratio is followed vertically up the graph until it intersects with aline that corresponds to the inlet pressure. The y axis value forthe point of intersection is the value for the loss correction

factor. For nonlubricated reciprocating compressors, the losscorrection factor should be multiplied by two. If the alternatevolumetric efficiency equation is used, the loss correction factorfrom Figure 7 must be multiplied by 100.

Figure 7. Loss Correction Factor for Reciprocating Compressors

The volumetric efficiency equations can be used during aprojects estimating phase to approximate the actual volumetricefficiency that is quoted by the vendor.


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CylinderDisplacement

Calculation of cylinder displacement can be used with themeasurement of actual cylinder volume to determine thevolumetric efficiency of any positive-displacement compressor inthe field. The actual capacity that is measured in the field isequal to the product of the displacement and the volumetricefficiency. The cylinder displacement of a reciprocatingcompressor is calculated through use of the following equations:

Single-acting compressor:

1728

nLmA

D

s

=

Double-acting compressor (without tail rod):

( )1728

nLma2AD s

=

Where:

D = Displacement, ACFM (actual cubic feet/minute)

A = Cross-sectional area of cylinder, sq. in.

a = Cross-sectional area of piston rod, sq. in.

m = Number of cylinders (for each stage)

Ls = Length of stroke, in.

n = Speed, strokes/minute, or rpm of crankshaft

The calculated displacement can be used to determine theactual volumetric efficiency as follows:

/hr)m(ACFMntdisplaceme

/hr)mor(ACFMcapacitymeasuredactualVE 3

3

=

When the actual volumetric efficiency equation is used, bothvolumetric flow rates must be in the same units (ACFM orm3/hr).


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Percent Clearance

The percent clearance for each stage is typically provided by

the compressors manufacturer. The percent clearance for astage can be determined through use of the following equation:

( )100innt,displacemepiston

involume,clearanceC

3

3

=

Calculating Discharge Temperature

The effects of discharge temperature on a reciprocatingcompressor encompass deterioration of packing, carbonization

of oils, and combustibility of gases. Packing life may besignificantly shortened by the dual requirement to seal both highpressure and high temperature gases. To reduce carbonizationof the oil and the danger of fires, a safe operating limit fordischarge temperatures may be considered to be approximately

300F. When handling gases containing oxygen, which couldsupport combustion, there is a possibility of fire and explosionbecause of the oil vapors that are present.

There are limiting factors when considering dischargetemperature. The maximum ratio of compression that is

permissible in one stage is usually limited by the dischargetemperature, particularly in the first stage. There are certainprocesses that require a controlled discharge temperature. Forexample, the compression of gases, such as oxygen, chlorine,and acetylene, requires that the temperature be maintained

below 200F, but for most field applications, the use of 300Fmaximum is a good average, and this maximum isrecommended by the API for nonlubricated compressors;however, in accordance with 31-SAMSS-002, compressorvendors are to determine the safe maximum dischargetemperature.

In accordance with SAES-K-403, the following guidelines apply:When compressors supply an air dryer, the maximum allowable

discharge temperature downstream of the after cooler is 50C(120F) for silica gel desiccant, and 60C (140F) for activatedalumina desiccant. Compressors that are equipped with

refrigerated dryers are limited to a 40C (100F) air dischargetemperature. Compressors in hydrogen service are limited to


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135C (275F) discharge temperature.

Calculation of compressor discharge temperature may beestimated as the adiabatic temperature at the end of the

compression process, which is given by the following equation:

Td = Tsx (R(k-1)/k

)Where:

Td = Discharge temperature, R

Ts = Suction temperature, R

R = Pressure ratio cylinder (discharge pressuredivided by suction pressure, psia)


The actual discharge temperature will differ from the calculatedvalue because the equation does not include the effects ofcylinder cooling and efficiency.

Calculating Power

The manufacturers performance predictions should always be

used as a first choice in the calculation of the power of acompressor. If the performance predictions are not available,the equations that are provided below can be used to makereasonable approximations. The equations to calculateisentropic horsepower are as follows:

1k

kQ43.67HP

= (R(k-1)/k-1)

Where:

HP = Isentropic horsepowerQ = Compressor capacity in million standard cubic

feet per day (mmscfd) at 14.7 psia and 520R


k = Isentropic volume exponent at operating


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conditions (the specific heat ratio for ideal gas,Cp/Cv)

The isentropic horsepower equation indicates that the major

effects on isentropic power are the capacity (Q) and thepressure ratio (R).

Any factors that affect capacity - for example, valve losses andleakage - increase the horsepower that is actually consumed bya compressor. To compensate for these factors, the isentropichorsepower equation can be modified by factoring thecompressors efficiency and its mechanical efficiency. Thefollowing equation provides a more accurate calculation of acompressors brake horsepower:

1kkxQx43.67bhp

= (R(k-1)/k-1) x

Where:

bhp = Brake horsepower

Q = Compressor capacity in million standard cubic

feet per day (mmscfd) at 14.7 psia and 520R



c = Compression efficiency

m = Mechanical efficiency

The compression and mechanical efficiencies can be suppliedby the compressors manufacturer. Compression efficiencyvaries with many factors, but the typical industry standard is

0.85 (85%) for lubricated double-acting reciprocatingcompressors and 0.80 (80%) for nonlubricated and/or single-acting reciprocating compressors. The typical industry standardfor mechanical efficiency is 0.95 (95%).

Screw compressor compression efficiency is typically 80% forlubricated compressors and 75% for nonlubricatedcompressors.


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WORK AIDS

WORK AID 1: RESOURCES USED TO DETERMINE DYNAMIC

COMPRESSOR ACCEPTABILITY

Work Aid 1A: Calculation Procedures

Convert the inlet and discharge pressure to psia.

P1= ( ) + ( ) = ( ) psia

P2= ( ) + ( ) = ( ) psia

Convert the inlet and discharge temperature to R.

T1= ( ) + ( ) = ( )R

T2= ( ) + ( ) = ( )R

The polytropic head is given by:

1-

P

P

1-n

nRTZ=H

1

2

1)/n-(n

1avgp


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To calculate Zavg:

From Table 1in Work Aid 1B

.

Pc= ( ) psia

Tc= ( ) R

At the inlet

)()(

)(==

P

P=P

c

1

r

)(=)(

)(=

T

T=T

c

1r

)(=Z1 from Figure 9 in Work Aid 1B

At the outlet

)(=)(

)(=P

P=Pr

c

2

)(=)(

)(=

T

T=Tr

c

2


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9Figurefrom)(=Z2 in Work Aid 1B

)(=2

)(+)(=

2

Z+Z=Z

21avg

To calculate (n-1)/n

P

P=

T

T

1

2

1)/n-(n

1

2

=

(n-1)/ n

( )

( )

( )

( )

)(=)(ln

)(ln=

n

1)-(n


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To calculate gas constant (R)

)(=)(

)(

=MW

R

=Runiv

Refer to Table 1 for the molecular weight of ethane.

The polytropic head equation becomes

ft)(=

1-)(

)(

)(

)())()((=

1P

P

1n

nRTZH

)(

1)/n(n

1

2avgp

=

To calculate the inlet flow (ACFM)

)(=

)(520

)(

)(

14.7)(=

Z520

T

P

14.7SCFM=ACFM 1

1

1


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To calculate polytropic efficiency

)(=)(

)1)/(-(

=1)/n-(n

1)/k-(k

=p

To calculate the gas horsepower

( )

hp)(=

)(33000)(

))((=

(33000)

lb/minRateFlowx)(H=GHP

p

To calculate the brake horsepower

( ) ( )

bhp)(=

=

lossesmechanicalghp=bhp

+

+


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Work Aid 1B: Pertinent Data

Nomenclature

Hp = Polytropic head


R = Gas constant


n = Polytropic exponent

P1 = Inlet pressure


Runiv = Universal gas constant

MWgas = Molecular weight of a gas

p = Polytropic efficiency

k = Isotropic exponent

Cp = Specific heat at constant pressure

Cv = Specific heat at constant volume

r = Compression ratio

ACFM = Actual cubic feet per minute

SCFM = Standard cubic feet per minute

Z = Compressibility factor

MCp = Molar specific heat at constant pressure

MCv = Molar specific heat at constant volume


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Charts for DeterminingCompressor PerformanceCharacteristics

Table 1. Critical Constants of Gases

Critical Constants

Compound FormulaMol. Wt.

MPressurepsia Pc

Temp.

R TcAcetylene C2H2 26.036 905.0 557.4

Air N+O2 28.966 547.0 238.7

Ammonia NH3 17.032 1,657.0 731.4

Benzene C6H6 78.108 714.0 1,013.01,2-Butadiene C4H6 54.088 653.0 799.0

1,3-Butadiene C4H6 54.088 628.0 766.0

N-Butane C4H10 58.120 550.7 765.6

Isobutane C4H10 58.120 529.1 734.9

N-Butene C4H6 56.104 583.0 755.6

Isobutene C4H6 56.104 579.8 752.5

Butylene C4H6 56.104 583.0 755.6

Carbon dioxide CO2 44.010 1,073.0 548.0

Carbon monoxide CO 28.010 510.0 242.0

Chlorine Cl2 70.914 1,120.0 751.0

Ethane C2H4 30.068 708.3 550.1Ethyl alcohol C2H5OH 46.069 927.0 629.6

Ethylene C2H4 28.052 742.1 509.8

N-Hexane C6H14 86.172 439.7 914.5

Helium He 4.003 480.0 510.0

Hydrogen H2 2.016 188.0 60.2

Hydrogen sulfide H2S 34.076 1,306 672.7

Methane CH4 16.042 673.1 343.5

Methyl alcohol CH3OH 32.042 1,157.0 924.0

Nitrogen N2 28.016 492.0 227.2

N-Octane C8H18 114.224 362.1 1,025.2

Oxygen O2 32.00 730 278.2

N-Pentane C5H12 72.146 489.5 845.9Isopentane C5H12 72.146 483.0 830.0

Propane C3H8 44.094 617.4 666.2

Propylene C3H6 42.078 667 657.4

Sulfur dioxide SO2 64.060 1.142 775.0

Toulene C7H8 92.134 611 1,069.5

Water H2O 18.016 3,206 1,165.4

Hydrogen chloride HCl 36.465 1,199.2 584.5


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Figure 8. Compressibility Factors at Low Reduced Pressure


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WORK AID 2: RESOURCES USED TO DETERMINE POSITIVE-DISPLACEMENT COMPRESSOR ACCEPTABILITY

Work Aid 2A: Calculation Procedures

Convert the suction and discharge pressure to psia.

P1= ( ) + ( ) = ( ) psia

P2= ( ) + ( ) = ( ) psia

Calculate the compression ratio:

R =

1

2

P

PR=

( )( )

( )=

Convert the inlet temperature to R.

T1= ( ) + ( ) = ( )R


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Calculate the volumetric efficiency, using Figure 9 in Work Aid 2B.

VE = 100 - R- L - C

1R

Z

Z 1/k

d

s

( ) ( ) ( ) ( )( )

( )( )

( )=

=

VE

1100VE1

Calculate the compressor displacement.

( )1728

nLma2ADISP s

=

( ) ( )( ) ( ) ( ) ( )1728

DISP

=2

( )=DISP


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Calculate the compressor capacity.

VEDISPZ

Z

T

P0.0509Q

s

std

s

s =

( )( )

( )( )

( ) ( )= 0.0509Q

( )Q=

Calculate the compressor brake horsepower.

( )mc

1/kk

1x

1x1R

1k

kxQx43.67bhp

=

( ) ( )( )

( )( )

( )( ) ( ) ( )111143.671

=

( )=


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Work Aid 2B: Pertinent Data

Hydrogen Sulfide

Zs........................ 1.015

Zstd...................... 0.99

Zd........................ 1.022

k........................ 1.320

Nomenclature

A = Cross-sectional area of cylinder, sq. in.

a = Cross-sectional area of piston rod, sq. in.

bhp = Brake horsepowerC = percent clearance as a decimal fraction of displaced volume

D = Displacement, ACFM (actual cubic feet/minute)

DISP = Cylinder displacement in cubic feet per minute (cfm)

HP = Isentropic horsepower

k = Isentropic volume exponent at operating conditions (the specificheat ratio for ideal gas, Cp/Cv)

L = Loss correction factor

Ls = Length of stroke, in.m = Number of cylinders (for each stage)

n = Speed, strokes/minute, or rpm of crankshaft

c = Compression efficiency

m = Mechanical efficiency

Pd = Discharge pressure in psia

Ps = Suction pressure in psia

Q = Capacity in million standard cubic feet per day (mmscfd) at 14.7 psia

and 520RR = Pressure ratio across the cylinder (discharge pressure divided by

suction pressure, psia)

Td = Discharge temperature, R

Ts = Suction temperature in R

VE = Volumetric efficiency


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Zd = Compressibility factor at discharge conditions

Zs = Compressibility factor at suction conditions

Zstd = Compressibility factor at standard conditions

Figure 9. Loss Correction Factor for Reciprocating Compressor


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GLOSSARY

adiabatic

compression

A compression process in which no heat is added or removed.

best efficiencypoint (BEP)

The point on the performance curve of a centrifugal compressorat which the efficiency is at a maximum.

brake horsepower(bhp)

The total horsepower that is required to drive a compressor.Brake horsepower is equal to the sum of the gas horsepowerand the mechanical losses.

compressibilityfactor, Z

The actual volume of a gas divided by the volume of the sameweight of ideal gas at the same molecular wei

Documents

Determining Compressor Acceptability