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The design, selection and applicationof oil-free screw compressorsfor fuel gas service

Klaus D. Lelgemann


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The Society shall not be responsiblefor statements of opinions advancedin papers or discussion at meetings of the Society or of its Division or Sec-tions, or printed in its publications.

Discussion is printed only if the paperis published in an ASME Journal.

IGTI-Vol. 8, ASME COGEN-TURBO, ASME 1993

Introduction

The design, selection and

application of oil-freescrew compressors forfuel gas service

Klaus D. LelgemannMAN Turbomaschinen AGGHH BORSIGGermany

Fuel gas compressors installed in co-generation systems must be highly reli-able and efficient machines. The screwcompressor can usually be designed tomeet most of the gas flow rates andpressure conditions generally requiredfor such installations.

To an ever-increasing degree, alter-native sources are being found for thefuel gas supply, such as coke-ovengas, blast-furnace gas, flare gas, land-fill gas and synthesis gas from coalgasification or from pyrolysis.

A feature of the oil-free screw com-pressor when such gases are beingconsidered is the isolation of the gascompression space from the bearingand gear lubrication system by usingpositive shaft seals.

This ensures that the process gas can-not be contaminated by the lubricatingoil, and that there is no risk of loss of lubricant viscosity by gas solution inthe oil.

This feature enables the compressedgas to contain relatively high levels of particulated contamination withoutdanger of "sludge" formation, and also

permits the injection of water or liquidsolvents into the compression space,to reduce the temperature rise due tothe heat of compression, or to "wash"any particulate matter through the com-pressor.

Natural gas is usually the most com-mon fuel used in thermal energy pro-cesses. However, alternative gases arenow being increasingly employed, suchas coke-oven gas, blast furnace gas,flare gas, landfill gas and synthesis gasobtained from coal gasification or frompyrolysis.

These gases are usually not as easyto process as natural gas and presentspecial problems to most compressordesigns.

The principle of operation of the screwcompressor has been known since1878, the year in which the respectivepatent was issued to Heinrich Kriegar.

The first operational compressor of this type was designed and construc-ted in 1934 by the Swedish engineer

Lysholm.

SRM (Svenska Rotor Maskiner AB), theowner of the master patents, has deve-loped this product to its present level of marketability in co-operation with licen-sed manufacturers throughout theworld.

Contents

1) Operating principle2) Design principle3) Tip speed4) Volume flow control by speed

variation5) Oil-free screw compressors with

liquid injection6) Mechanical features of the oil-free

screw compressor7) Shaft seals8) Typical screw compressor sizes

and performance data9) Examples of application and design

2

Abstract The American Society Of Mechanical Engineers345 E, 47th St., New York,N.Y. 10017


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A screw compressor comprises tworotors one of which is called the maleand the other, the female rotor. Eachrotor has a number of profiled lobeswhich are designed to mesh with eachother. The male rotor has a similarfunction to that of a piston in a recipro-cating compressor and provides induc-tion, compression and discharge bymeans of purely rotary motion.

This means that the operating principleof the screw compressor is that of apositive-displacement machine.

Consequently there is no danger of sur-ging as in the turbocompressor. Themachine has a stable characteristicwith a rise in the pressure ratio hardlycausing any drop in volume flow. It has

no valves that could become defective,and no component parts subject tomixed friction and wear.

Thanks to the purely rotary motion nofree inertial forces are produced whichwould necessitate special foundationsor could excite vibration of neighbou-ring components.

The machine has, in relation to its

weight and dimensions, a high specificperformance.

The induction, compression and disch-arge operations can only be performedwhen the rotors are housed in a casingwhich allows proper admission to therotor profiles through an aerodynami-cally shaped inlet port and is providedwith an outlet port of similar design.

The inlet port is at the top and the out-let port at the bottom. The requiredpressure ratio determines the built-involume ratio (Vi) for a particular fuelgas, and this inturn governs the sizeand configuration of the discharge port.

The volume ratio determines theamount of compression work prior todischarge.

Since the rotors in an oil-free screwcompressor are not in contact witheach other, some means of keepingthem separated and in phase is requi-red. This is achieved by means of a pairof timing gears usually adjusted withzero backlash to ensure that the designtolerances are maintained and therotors are properly synchronized whenrunning. Only about 10% of the powerinput to the male rotor is transmitted tothe female rotor. The male rotor, whoselobes mesh with the flutes of the fema-le rotor, causes displacement and ejec-tion of the gas.

The bearings ensure that an exactlydefined clearance is maintained bet-ween the rotor circumference and thecasing as well as between the end faceof the rotor profile body and the casing.

In order to achieve the highest possibleefficiency, the built-in volume ratioshould be as close as possible to thevolume ratio determined by the actualpressure ratio.

The rotors are radially and axially sup-ported in sturdy bearings outside thecompression space. Process-gasscrew compressors are supportedexclusively in sleeve-type radial bea-rings with a hydrodynamic film. Thesehave a life expectancy of a very highorder.

The rotor profile clearances are estab-lished by precise manufacture. Thetiming gears ensure that intermeshingof the rotors takes place without anycontact between the profiles.

Therefore, the rotors of the process-gas screw compressors do not needany lubrication in the compressionspace since there is no contact bet-ween the rotors. Consequently there isno danger of lubricant deterioration bygas or gas constituents or of any lubri-cant portions entering downstream sys-tems.

1 Rotor pair of a process gas screw compressor

2 Vertical cross-section of a typical "oil-free"process-gas screw compressor

2) Design principle(Fig. 2)1)Operating principle(Fig. 1)


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Screw compressors have a constanttorque at a constant pressure ratio.

This means that with a speed con-trolled screw compressor an excellent

partload response is obtained.

Admittedly, a negative effect is produ-ced by the previously described influ-ences, such as a drop in volumetric effi-ciency on reduction of the tip speeddue to a constant leakage backflow.

However, there is also a decrease inthe dynamic losses already referred to,so that proper control by speed varia-tion is possible within a relatively widerange. Here again the sonic velocity inthe gas is an important factor. Thegraph shows a screw compressor

without injection which handles a gaswith a molecular weight of 30 kg/ kmol.

The performance is such that operationaway from the optimum speed is bestachieved by means of variable-speeddrive.

The drop in volumetric efficiency at lowspeed causes the discharge tempera-ture of the compressor to rise.Heavy gases with a low sonic velocityallow speed variation within an evenwider range, as at low speed the vo-lumetric efficiency drops to lesserdegree.

With light gases the range of speedadjustment is limited for screw com-pressors without injection and allowsflow variation by about 30% only.

The oil-free screw compressor has al-ready been referred to several time asa machine without injection. It has also

been mentioned in the preface that thespecific features of the oil-free screwcompressor can still be improved byproviding additional facilities.

One important step for improvement isto inject liquids. This differs from thelubricant-injected screw compressor inthat the liquid injected is not expectedto lubricate the rotors and the bearingsat the same time and is recirculated ina closed-circuit oil system, but can beselected for the specific gas and theparticular requirements of the process.

5 Effect of water injection on discharge tem-perature for "oil-free" screw compressor

4 Curves showing effect of speed variation for"oil free" screw compressors

4) Volume flow control by speed variation (Fig. 4)

5) Oil-free screw com-pressors with liquidinjection (Fig. 5)


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B. Injection for cooling andwashing

Compression is known to be a more orless isentropic process. The compres-sion temperature results from the pres-sure ratio, the Cp/Cv factor of the gasto be compressed, and an allowancefor the loss in efficiency.

As the operating temperature of an oil-free screw compressor is restricted bythe selection of the construction ma-terials and the coefficients of expansionof these, i.e., their thermal expansion,the discharge temperature must belimited to about 200C or a maximumof 250C for each compressor stage.

The running clearances referred to in

section 2) may cease to exist if the dis-charge temperature is too high, andthis may result in damage to the ma-chine.

Liquid injection is very helpful where thegas to be compressed permanentlychanges its molecular weight and thusthe Cp/Cv factor.

An example of such gases is flare gas.

If the gas mixture is heavy, the Cp/Cvfactor is small; for instance, at a pres-sure ratio of 6 a single-stage compres-sor can be operated with a relativelylow discharge temperature.

If the gas mixture becomes lighter, theCp/Cv factor increases considerablyand the discharge temperature of thecompressor rises abruptly above theallowable level.

In this case a small quantity of liquid,such as petrol, is injected into the com-pressor stage and the vaporizationheat of this absorbs the heat of com

pression. The compressor can safelybe kept running.

Compressors for gases with changesin molecular weight between 9 and 44kg/kmol are already in existence.Other compatible liquids can also beused. It is only necessary to ensure thatvaporization takes place without anyresidues and that the liquid injecteddoes not cause any erosion or corrosi-on problems. The danger of erosionand corrosion can also be eliminatedby selection of suitable materials for thecompressor.

However, the gas may also necessitatelimitation of the compression tempera-ture. For instance, all acetylenecontaining mixed gases decompose inan explosion-like manner when no oxy-gen is present. Styrene containinggases polymerize and butadiene-

containing gases form rubber.

This example shows that injection of aliquid may be very important.

The quantities of liquid injected for tem-perature reduction are relatively smallcompared to the weight of the com-pressed gas, as the heat is removedwith the aid of the vaporization heat of the liquid.

This thermal limitation allows single-stage compressors to be applied incases where normally two-stagemachines would be needed.In addition, variable-speed control ispossible within a wide range.

reduction of the compression tempera-ture to prevent polymerization or reac-tions in the gas handled. The compres-sion process is almost isothermal.

In addition, the liquid present in thegaps also causes a slight in improve-ment in volumetric efficiency. There isalso a washing effect.

Considerable quantities of partly unde-sirable constituents of the gas aretransferred to the injected liquid if theundesirable constituents have an affini-ty to the liquid. Consequently, thereluctance to use these gases with gasturbines is rapidly diminishing.

Among these is coke-oven gas, theNH3, H 2S and HCN contents of whichare considerably reduced.

Another example is low-caloric furnace

top gas, where the contents of chlorideand also of dust are reduced. There areoil-free screw compressors which hand-le gas containing up to 300 mg/m 3 of dust without any noticeable abrasioneffect. This is explained by the fact thatthe liquid injected covers all internal sur-faces of the machine with a film bywhich the dust is kept off.

However, the material selected for the

compressor must be of a kind whichdoes not allow erosion to be caused bythe medium injected.

The injection of large quantities of liquidinto oil-free screw compressors alsoallows landfill gas to be handled bythese machines. With this gas onenever knows for sure what aggressiveconstituents and in what quantitieshave been added to the base gas inthe course of time. As the injectionliquid is incompressible, it is necessaryto limit its quantity according to therotor profile chosen and to the built-involume ratio.

If a quantity of liquid greater than thevaporization quantity is injected, a two-phase mixture is discharged at the com-pressor outlet port. This liquid causes a

A. Injection for reductionof the compression tempe-rature

One generally distinguishes betweentwo kinds of injection:


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Normally injection quantities up to 3%by volume referred to the suctionflowrate of the compressor stage donot present any problem if the liquid isdistributed in the form of droplets. Thisapplies particularly to machine wherethe liquid is already contained in thegas handled, as with two phasiccompressor.

Abrupt accumulation of large quantitiesof liquid may cause destruction of ascrew compressor (liquid hammer).

As the intake port is at the top and theoutlet port at the bottom, liquids auto-matically run out of the compressionspace when the machine is at a stand-still.

However, the intended cleaning of thegas stream from undesirable gas cons-tituents does not allow frequent utilisa-tion of the injection liquid in the com-

pressor. The contaminants accumulateand the liquid gradually loses its abilityto absorb them. A point of saturation isreached.

In addition, liquids with a high concen-tration of acid constituents start atta-cking the materials of the compressorsystem. Therefore it is necessary toremove injection liquid in good timeand to replace it by a clean liquid.

Some injection liquids are used onlyonce, which shows that the desiredcleaning of the main gas stream is notachieved without consequences, asthe liquid removed has to be discar-ded.

This compression process also allowssingle-stage compressors with highpressure ratios to be provided and tooperate these machines within widespeed variation ranges.

As the rotors are rotating bodies, therotor shaft stress is an alternatingstress. The ratio of actual stress toallowable stress has to be checkedin every particular case.

Depending on the material, a stress of up to about 40 N/mm 2 is allowed. Thisvalue is relatively low, but the notcheffect in the fillet between rotor profilebody and rotor shaft as well as any pos-sible corrosive attack on the shaft sur-faces by the gas have to be taken intoaccount. The bearing span results fromthe length of the profile body and thelength needed for shaft sealing.

The residual rotor shaft diameter resultsfrom the profile chosen is usually themost significant single factor affectingrotor deflection.

Compression takes place only in an ex-actly defined sealed triangular area of the rotor pair on one side, i.e., thedischarge side. This means that theopposite, or suction side, of the rotorpair is either exposed to no pressureor inlet suction pressure.

As a result, the rotor pair is subjectedboth radially and axially to forces pro-duced by the difference between thegas pressures. This is entirely differentfrom the turbocompressor.

These features lead to certain specificrequirements.

In the first place, rotor material of appropriate strength has to be chosen.

This is necessary because of the factthat the stress which arises and thedeflection of the rotors are determinedby the bearing spacing and the residualrotor shaft diameter together with thegas load. However, the small rotor pro-file clearances do not allow any majordeflection.

Small numbers of profile lobes, such aswith the 3/4 combination, result insmall residual rotor shaft diameters.

Large numbers of profile lobes, such aswith the 4/6 combination, result in lar-ger residual rotor shaft diameters andconsequently stiffer rotors.

6 Pressure distribution and rotor bending force fora downward discharge screw compressor

6) Mechanical featuresof the oil-free screwcompressor (Fig. 6)


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A typical length/diameter ratio of therotor profile body in an oil-free screwcompressor is 1.65, although modelswith an L/D ratio of 1.2 and 0.9 arealso available.

However, scope for development hasnot yet come to an end and the inno-vation potential is far from beingexhausted.

Larger lobe numbers and smaller L/Dratios are already under considerationwith a view to developing screw com-pressors for higher pressure differen-ces.

There are also other essential distingu-ishing features compared to turbocom-pressors. The necessary high rotor stiff-ness with respect to deflection and ro-tor shaft stress automatically leads tolateral critical speeds which are above

the operating speeds.

The separation margin of 20% over themaximum allowable operating speedas required under API 619, para4.7.1.6 is exceeded in most cases.

For a machine with variable-speedcontrol this means that operation atany desired speed within the adjustingrange is possible. Any necessary upra-

ting later on can also be achieved inmost cases by just increasing thespeed.

The bearings have to absorb the radialand axial forces resulting from the dif-ference between the gas pressures.

These forces are much higher thanthose arising with turbocompressors,where generally only the rotor weighthas to be supported. Therefore the bea-rings of screw compressors are provi-ded with much larger faces.

The allowable load for radial bearingsis max. 5 N/mm 2 and for axial bearingsmax. 2 N/mm 2.

The relatively high bearing load in ascrew compressor is always applied inthe same direction. Therefore no oilwhip can occur as in the turbocom-pressor, where this is prevented by pro-viding bearings of special shape (poly-gon shape, lemon shape, tilting pad).

The bearings used by MAN TURBO forscrew compressors are sleeve-typeradial bearings and tapered-land axialbearings.

While with turbocompressors, the me-chanical bearing loss is about 1% of the compressor power requirement,this loss may be up to 6% in the caseof a screw compressor.

Of course, the higher power losses arepartly due to the fact that two rotorshave to be supported and there aretwice as many bearings.

This disadvantage does however provi-de a desirable beneficial feature.

Thanks to the previously mentionedhigh rotor stiffness, a rigid connectionbetween the rotating rotors and the stiff casing is established with the aid of thehigh bearing forces. This allows any irre-gular condition with respect to unba-lance to be exactly measured with anaccelerometer to API 670.

As the machine is run at subcriticalspeed, dynamic balancing of the rotorsis never done in the contract casing, sothat there is no need for vibration mea-surement in a second plane or formounting a key-phasor.

Of course, installation of vibration pro-bes to API 670 is possible. However,these do not provide any more safetyof compressor operation.

For measurement of the bearing condi-tion, RTD's to API 670 are providedboth radially and axially in the load-supporting areas of the bearing faces.

As the rotors intermesh in the profilebody and there are only small clearan-ces, an RTD embedded in the axial bea-ring responds to an axial rotor move-ment due to a bearing defect morequickly and with more reliability than anaxial position probe to API 670.

Therefore, installation of axial positionprobes to API 670 can be dispensedwith.

Of course, installation of axial positionprobes to API 670 is possible. How-ever, these do not provide any moresafety of compressor operation as theyonly respond when an axial bearing isdefective and the compressor hasalready been shut down by the axial-bearing temperature monitoring sys-tem.

For radial vibration measurement it isalso important to know that the lobes

on the male rotor cause a radial vibrat-ion. Each lobe meshing with the corre-sponding flute on the female rotorcausing a pulse, like gas ejection per-forms an operation which correspondsto that of the compression stroke of areciprocating compressor.

This compression stroke of the male-rotor lobe produces a reaction vibrationin a radial direction on the rotor and thecasing.

This reaction vibration caused by thegas pulsation is insignificant for a quali-tative assessment of the runningsmoothness of the machine to VDI2056 "Criteria for assessing mechanicalvibrations of machines" and has there-fore to be filtered out at the monitor.

This also applies to low-frequencyvibrations as otherwise vibrations pro-duced in the vicinity of the compressormay cause compressor shutdown.

The vibration amplitude produced bythe compression stroke of each rotorlobe may be up to 4 times that produ-ced by the rotation of the male rotor.


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8) Typical screw com-pressor sizes andperformance data

A portable real-time analyser allowsoccasional checking of the entirevibration spectrum.

The relatively high bearing losses in ascrew compressor also have a furtheradvantage which should not remainunmentioned. This is the high dampingeffect with respect to torsional criticalspeeds, because in drivers with a relati-vely uniform torque, such as squirrel-cage motors, steam or gas turbines,the damping forces in the rotating sys-tem are so high that even torsional for-ces excited at the first torsional criticalspeed cannot produce surface stres-ses that are a danger to the system.

This is particularly important with varia-ble-speed control ranges between 50and 100%, where the first torsional cri-tical speed is frequently passed.

Of course, for drivers with an irregular

torque the transient torsional vibrationanalysis required under API 619, para.4.7.2.6, is performed.

Most "oil-free" screw compressors areequipped with mechanical (contact)seals. These are cooled by the oil sys-tem of the compressor. With this typeof sealing, buffer gas is needed in veryrare cases only.

However, buffer gas can be additionallyprovided where difficult or toxic gasesare compressed. This seal is gas-tightnot only during operation of the machi-ne but also during standstill periods,even if the seal oil supply should fail.Leakage oil is generally discarded.

With currently available material combi-nations for mechanical seals, such assilicon carbide, mechanical seals are nolonger regarded as wearing parts.

All "oil-free" screw compressors arealso available with double or single drygas seals. The most appropriate sea-ling system is agreed upon between

the user and the manufacturer.

It is necessary to provide an absolutelytight sealing between the gas-handlingportion of the machine and the lubrica-ted hydrodynamic bearings, as no con-tact between the gas and the lubricantis allowed.

This is the only way to achieve the longinspection intervals of 24,000 servicehours for the compressor units, de-manded today for applications in refi-neries, power generation and petro-chemical plants, and to obtain a highavailability.

Labyrinth seals in a variety of materialscan be provided. These seals require abuffer gas or gas extraction system.Most of the oil-free screw compressorswith liquid injection as described under5. B. are provided with liquid-film seals

whose sealing liquid is the same as themedium injected into the machine.

7) Shaft seals


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Fig. 7

Fig. 8 Fig. 9

Process-gas screw compressor com-pletely assembled at the manufactu-rer's works as a packaged unit forinstallation directly at a well-head (gasgathering) without needing a foundati-on. The unit is protected againstingress of drifting sand. Cooling of thelube and seal oil for the mechanicalseal is by an air-blast cooler as the com-pressor is used in a desert area. It com-presses 3,950 m 3 /h of "associate" gasfrom 1.03 bar to 3.44 bar.

Flare-gas compressor with accessiblenoise enclosure in a refinery.

Two-stage process-gas screw com-pressors are mounted on a single tab-le-top foundation. All auxiliary equip-ment is installed under the table-top.Each individual compressor is equip-ped with liquid-film seals.

The same liquid which is used for sea-ling is injected into the compressionspaces to prevent polymerisation of theheavy hydrocarbons. Compression isalmost isothermal. Each unit compres-ses 11,000 m 3 /h of flare gas from 1.01to 7.0 bar.

Fuel-gas compressors in a refinery. Three two-stage process-gas screwcompressors (two driven by electricmotor, one with steam turbine drive) aspackaged units, mounted on concreteslabs, for fuel gas supply to a gasturbine.

The compressors are equipped withmechanical seals. Each unit compres-ses 2,500 m 3 /h of fuel gas from 3.5 barto 17.5 bar.

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9) Examples of application and design


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Fig. 10

Figs. 11 and 12 Fig. 13

Charging a molecular sieve with coke-oven gas for hydrogen extraction.

Two two-stage process-gas screw com-pressors as packaged units mountedon concrete slabs, outdoors.Each individual compressor is equip-ped with liquid-film seals. The sameliquid which is used for sealing is injec-ted into the compression spaces to pre-vent polymerisation of the heavy hydro-carbons. Compression is almost iso-thermal. Each unit compresses 4,800m 3 /h of coke-oven gas from 1.03 barto 9.0 bar.

Process-gas screw compressor com-pletely assembled at the manufactu-rer's works as a single-lift package for

direct mounting on the deck of an off-shore oil production platform. Themachine compresses 5,650 m 3 /h of "associate" gas from 1.37 to 6.85 bar.

This compressor stage has mechanicalseals.

Injection of coke-oven gas into blast fur-naces to substitute coke and to acce-lerate reduction. Two two-stage pro-

cess-gas screw compressors are indi-vidually mounted on table-top founda-tions. All auxiliary equipment is installedunder the table tops. The gas/watercoolers and the water/air coolers arelocated adjacent to the foundations.Each individual compressor is equip-ped with liquid-film seals. The sameliquid which is used for sealing is injec-ted into the compression spaces toprevent polymerisation of the heavyhydrocarbons. Compression is almostiso-thermal. Each unit compresses40,700m 3 /h of coke-oven gas from1.01 bar to 11.0 bar.

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MAN Turbomaschinen AGGHH BORSIGSteinbrinkstrasse 146145 Oberhausen / GermanyPhone +49. 208. 6 92-01Fax +49. 208. 6 92-20 19www.manturbo.com

MAN Turbomaschinen AGGHH BORSIGEgellsstrasse 2113507 Berlin / GermanyPhone +49. 30. 440 402 0

Fax +49. 30. 440 402 2000

MAN Turbomaschinen AGSchweizHardstrasse 3198023 Zurich / SwitzerlandPhone +41. 1. 278-22 11Fax +41. 1. 278-29 89

MAN Turbomacchine S.r.l.De Pretto

Via Daniele Manin 16/1836015 Schio (VI) / ItalyPhone +39. 0445. 6 91-5 11Fax +39. 0445. 5 11-1 38

In the interests of technical progress,subject to revision without notice.Printed in Germany.May 2002

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