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Suite 9 -2611, 37h Avenue NECalgary, AlbertaT1Y 5V7Call: (403) 250-9220Website: www.detechtion.com

C'IAND FLEET ENT EiPEEIALIS.I

Utilizing the

Rotary Screw Compressor

Compiled By: Brian TaylorDirector of Engineering

Copyright 2002Revision 1.0

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resston:

Table of Contents

Page. 1.0 The Physical Properties of Natural Gas.

1.1 Basic Gas Laws1.2 Atmospheric Pressure1.3 Elevation1.4 Abbreviations for Pressures1.5 Gauge Pressure'1.6 Absolute Pressure1.7 Boyle's Law1.8 Absolute Temperature1.9 Charles' Law1.10ldealGas Law1.1 1 Compressibility Factor1.12 Specific Heat Ratios1.13 Water Content of Natural Gas

4

2.0 The Compressor Hardware

2.0 General Description2.'1 Rotary Screw Compressor

2.1.1 Frames2.1.2 Rotors2.1.3 Balance Piston2.1.4 Thrust Bearings2.'t.5 Radial Bearings

2.2 Auxiliary Hardware2.2.1 lnlet Separator2.2.2Oi4 Separator2.2.3Oa1Cooler2.2.4Gas Cooler2.2.5 NormalGas Flow

3.0 The Compression Process

3.1 The Compression Process3.2 The Gompression Cycle

17

28

3.3 Estimating Compressor Capacity3.4 Compressor Displacement3.5 Volumetric Efficiency3.6 Capacity Contro!

3.6.1 Suction valve Control3.6.2 Variable Speed3.6.3 Slide Valve

3. 7 Compressor Horsepower3.7.1 Theoretical Horsepower3.7.2 Brake Horsepower3.7.3 Horsepower Calculation Example

3.8 Adiabatic Efficiency

2 E€'T.€EHTTETFI

2.0 General Description

Rotary screw compressors have been applied to various differentapplications since their commercial use began more than 50 yearsago. These machines can be found most commonly in air serviceand refrigeration duty. However, the screw compressor isbecoming more and more of an integral part of the natural gasprocessing industry.

The screw compressor is a positive displacement, volume reductionmachine, and can be either oil free or oil flooded in design. The oilfree packages are used mainly in "dirty" gas applications, flare gasrecovery, etc. and rely on timing gears to synchronize both therotors. The process gas areas are sealed off from the lubricationsystem and bearing cavities by means of adequate sealingtechnology. The oil flooded screw compressor is used in air andprocess gas applications and relies on oil injection directly with thecommodity that is being compressed. The injection of oil and its'importance in the compressor wil! be discussed later in this paper.

The basic design of the screw compressor (see Figure 2.1) consistsof two helical rotors that are cut in overlapping spirals within acommon housing. The two rotors are identified as being a male or afemale rotor and are supported by thrust and radia! bearings. Themale rotor is driven and the female rotor is the idling one. This willbe explained in further detai! later in this manual. A goodunderstanding of the basics of operation will help to avoid problemsand allow an individual to apply the compressors correctly todifferent industrial applications.

There are many benefits that are offered by the use of a rotaryscrew compressor which include the following:

a. Low operating and maintenance costsb. Low purchase pricec. Ability to handle low suction pressures - down to 26

inches of vacuumd. High compression ratios - up to 16e. Very portable - can be trailer mounted

18 E€"Y-WSffi EH TIEI-I

Suction Flange

2.1 Rotary Screw Compressor

Male Rotor

Balance Piston Female Rotor

-

Drive Shaft

OilGalleries

=j

Frame

Bearings

TimingGears

Figure 2.1 - Screw Compressor Cut-Aw?yr (courtesy of Mycom)

2.1.1 Frames

The cast iron frame houses all of the compressor's internal partsand distributes the oil through the various oil galleries in the case.The frame provides central support for the bearings, which in turnsupport the rotors. The main stress on the frame structure is thebuild up of internal pressures and heat loads during thecompression process. Maximum discharge pressures ofapproximately 300-350 PSIG are normal for most manufacturers.

19 EI€-T€EH'.r.IEfA

2.1.2 Rotors

Two major components of the screw compressor are the helica!rotors necessary for compression. They are designated as the malerotor and the female rotor. The male rotor is the driven rotor.meaning that it is connected to the driver either directly or througha gearbox. The male rotor typically contains four or five lobes thatmesh with the female rotor. The female rotor is the idling rotor androtates based on the speed of the male rotor. The female rotortypically contains six or seven "inter-lobes" that compliment themale lobes to create the cavity for the gas to fill. lt is important tonote that there isn't any meta!-on-metal contact between the maleand female rotors. A very thin oil film prevents the two rotors fromtouching as they rotate in an oil flooded application and timinggears prevent contact in the oil free design.

The rotors are machined to a very high tolerance. Whenassembled, the gap between the lobes as they mesh during rotationis less than 0.005 inches. They are also dynamically balanced whenthey are manufactured to prevent vibration at high speeds.

The sizes of the rotors that are found in different models ofcompressors help dictate the performance that the compressor iscapable of. A large diameter rotor is capable of moving largervolumes of gas than a smaller diameter rotor. At the same time, along rotor is capable of moving a volume of gas through a higher-pressure ratio than a shorter one.

The profile of the rotor plays a very significant part in theperformance of the machine. There is leakage that occurs betweenthe rotors, between the rotors and housing bore surfaces and at theends of the rotors and housing. One major leakage path is knownas the "blow hole". The blow hole area is a triangular cusp that isformed between rotor lobe surfaces and the housing bore. Going toan asymmetric lobe profile design has minimized this blow holepathway and increases efficiencies. Figure 2.2 shows the twodifferent lobe profiles, asymmetric and symmetric, andconfigurations possible within a screw compressor.

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Figure 2.2 - Rotor Profiles

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2.1.3 Balance Piston

The balance piston is a hydraulically controlled disc or piston thathas been designed to offset the thrust force on the thrust bearings.!t exerts a force on the male rotor, since, by its' design, it receives agreater thrust force than what the female rotor receives. Thebalance piston effectively opposes the thrust load due tocompression, and it exerts force based on discharge gas pressures.Itis veryimportant to keep an adequate amount of oil pressure onthe balance piston so that the thrust bearings and/or roller bearingsdon't fail prematurely. At the same time, too much pressure on thebalance piston can also have adverse affects, such as overpressurizing and causing a reverse thrust force. Again, thissituation can also result in premature bearing failure.

2.1.4 Thrust Bearings

The thrust bearings are what prevent the rotors from contactingeach other and the inside of the compressor housing. They are alsoresponsible for absorbing the axial thrust that is generated by thegas being compressed. The thrust bearings consist of a twobearing assembly, each of which absorbs forces in opposite axialdirections, ie. suction to discharge pressures OR discharge tosuction pressures.

2.1.5 Radial Bearings

The radial bearings carry the weight of the rotors plus the radialforce produced by the gas pressure. The bearings used to carrythis radial load are known as hydrodynamic or frictionless bearings.The shafts of the rotors "float" in a pressurized oil film within thebearing housing. This allows for an infinite L10 bearing life if the oilsystem and oil quality are kept in good check.

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2.2 Auxiliary Hardware

2.2.1 lnlet Separator

The inlet separator is a vertical vessel that is mounted to the skid infront of the inlet to the compressor. This vessel acts as a filteringelement for the incoming gas, in order to protect the compressorfrom both solid and liquid debris. Due to the very tight tolerancesand high speed of the compressor components, this piece ofequipment is necessary to prevent any foreign material fromentering the compressor and damaging the rotating parts.

The principles employed in removing materialthat is trapped in thegas as it enters the separator are impingement, change of direction,change in velocity and a filter element also known as a wire-meshmist extractor.

As the gas stream enters the separator vessel, the gas is turneddownward to strike the impingement baffle. The heavy solids andliquids are then collected here and forced down through holes intoan accumulation chamber at the bottom of the vessel. This changein direction causes a decrease in velocity, and with the added helpof gravity, further dropout of solid and liquids continue. The gasstream then will pass through the mist extractor, where very fineparticles of solid and liquid that are still entrained in the gas stream,will coalesce on the filter. Once these droplets grow large enough,they wil! fall onto the impingement baffle. The clean gas can nowflow into the compressor.

The liquid that has been separated from the gas stream, and hascollected in the accumulation chamber, will be automaticallydumped once the laquid level controller senses high levels ofaccumulation. lt is important not to allow any of this Iiquid to enterthe compressor. ln the event that the dump valve cannot handle theremoval of all the liquid, a high liquid level shut down switch must beadded to prevent overflow into the compressor.

23 E€T{fiffi=HTIEN

2.2.2 Oil Separator

Although oil is an extremely important part of the compression cyclein an oil flooded screw compressor, the oil that mixes with the gasduring compression is not desirable after compression is complete.This warrants the need for an oil separator. lt is another skid-mounted vessel required by an oil flooded screw compressor toremove the oil droplets that become entrained in the gas duringcompression. See Figure 2.3 and 2.4lor an illustration of horizontaland vertical separators.

The principles used to knock the oil out of the gas are the sameones used in the inlet separator vessel as previously mentioned. ltis important to note that the gas/oil mixture must be kept at avelocity of less than 30ft/sec. This will prevent unwantedturbulence inside the separator. Turbulence inside the separatorcan cause unsteady oil levels, foaming, vibration and a lack ofseparation efficiency.

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Figure 2.3 - Horizontal Separator

Figure 2.4 - VerticalSeparator

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2.2.3 Oil Cooler

The primary role of the oil cooler is to cool the hot oil from thedischarge line from about 180'F back down to an acceptable re-injection temperature of about 140"F with as minimal a pressuredrop as possible. This is accomplished through the use of a shelland tube water-cooled arrangement or an aerial type exchanger.The installation of a 3-way mixing valve will allow the properregulation of oil temperature by mixing cold oil with hot oil beforere-iqjection. The oil is temperature mixed and filtered before re-entering the compressor. Figure 2.5 illustrates the flow of oilthrough a system such as the one described.

Figure 2.5 - Oil Cooler

25 E€TEEHTIETA

2.2.4 Gas Cooler

Heat is generated in a gas after it is compressed. Once the gas hasleft the oil separator, free of any oil droplets, it must be cooledbefore release into the final discharge line. The gas cooler works inthe same way that the oil cooler does. The cooler typically containsa fan that is either driven off the primary mover, or has its ownpower source such as a separate electric motor. The fan forces airover the cooler sections as the gas passes through them and theheat is extracted.

2.2.5 Normal Gas Flow

When dealing with a rotary screw compressor, there is typicallyonly one stage of compression present on a unit. The gas enters theinlet separator where all of the foreign !iquids and solids areknocked out of the gas stream. The gas then passes in to thesuction side of the compressor where it begins to be injected withoil. As compression begins, oil is continually injected into the gasstream. The compression process finishes when the gas/oil mixturereaches the discharge port. This gas/oil mixture enters the oi!separator where the gas is "cleaned" of all oil droplets. The gasthen leaves the separator, passes through the gas cooler and issent downstream in the discharge Iine. Figure 2.6 illustrates atypical gas compression system.

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3.0 THE COMPRESSION PROCESS

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3.1 The Compression Process

The compressor operates based on the fundamentals of Boyle's law- the pressure of a gas will increase when its volume is decreased.The compression process can be looked at from three similar, butdifferent relationships, when dealing with the screw compressor.The first and most basic relationship is:

ISOTHERMAL PROCESS

v2

V,l__t_

P2 (Eq.3.1)

This is a very simplistic but not a realistic view, in that it assumesconstant temperature throughout the cycle.

Another relationship that is commonly used to predict the behaviorof gas during compression is known as the Adiabatic Process. Thistheory takes into account the ratio of specific heat (k) of the gasundergoing a compression cycle, but ignores the transfer of heatbetween the gas and its surroundings and vice versa. The equationused for the adiabatic modeling of a gas is:

ADIABATIC PROCESS

PVn :P,Y, k

(Eq. 3.2)

29 ra€"?*-{ffi EHTtEtr-r

The third relationship that is most accurate in precisely defining thecompression of a gas in the "real world" is known as the PolytropicProcess. The following equation represents the Polytropicrelationship and uses the Polytropic exponent "n".

POLYTROPIC PROCESS

PV' -PrVrnk-1

where fl = (Eq. 3.3)kxrl

The Polytropic exponent (n) is derived from the specific heat ratiosof the gas (k) and takes into account the Polytropic efficiency (ry).The Polytropic efficiency, when dealing with a screw compressor,has a typical range of 0.85 to 0.95.

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Figure 3.1 below clearly illustrates how the three differentapproaches that define the compression cycle can differ.

tDL

=il?TEfl)L-0-

AB - AdiabatieAS " Folytrepir.AD - lsetherma$

FV=C

PV"=C

P\*'=E

A

Volurme

Figure 3.1 - PV Diagram

The area under each of the curves determines the amount of workthat has to go into the system to move the gas from suction todischarge pressure.

Work=PressurexVolume

Therefore, the larger the area under the curve, the more work isnecessary to compress the gas. lt can be seen then, that thelsothermal Process understates the horsepower required tocompress gas, whereas the Adiabatic Process overstates thehorsepower requirements. If one is to accurately predict thehorsepower expenditure for a set of gas conditions, the PolytropicProcess is the model to use.

31 E€'T-ffiffiHTIEI-I

3.2 The Compression Cycle

1 -SuctionSuction gas is drawn in at the TOP of thecompressor where the male lobe rotates out of thefemale inter-lobe in order to create an expandingcavity. A charge of fresh gas continues to fill theentire volume of each interlobe as the unmeshingthread proceeds down the length of the rotor. Thevolume of gas that fills the entire Iength of thread isknown as the suction volume (Vs).

View from BOTTOMof compressorI-I-II-III--

2 -TranslationRotation continues untilthe rotors have rotatedenough and moved the gas axially past the inlet portto sealthe suction charge from the inlet. The gas isnow trappedbetween the rotors and thecompressor housing. As the rotors continue torotate, they translate the gas axially towards thedischarge port. At this point the trapPed gas hasoccupied the lobes on the BOTTOM side of therotors.

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ressl0nThe rotors, continuing to rotate, have now moved thegas far enough down the length of the rotors, that thevolume occupied by the gas has now started todecrease. Because of the helical design androtational motion, this decrease in volume andincrease in pressure is gradual and smooth, therebyeliminating any pulsations or vibrations due to thecompression process.

4 -DischarThe process of compression is completed when thetrapped gas has reached the discharge port. Thelocation of the discharge port dictates the finalvolume of gas at discharge. The Ionger the gasremains trapped in the rotors (ie. the more axially thegas travels) the higher the internal pressuredeveloped and smaller the finalvolume. Likewise,the sooner the trapped gas reaches the dischargeport the lower the internal build up of pressure.Location of the discharge port is very important anddetermines the amount of under-compression orover-compression that will occur inside thecompressor.

33 E€TEEEHTIEIrl

3.3 Estimating Compressor Capacity

Compressor capacity is defined as the actual volume rate of flow(ACFM), or gas compressed and delivered, at the prevailingtemperatures and pressures at the compressor inlet. This capacityis converted to the standard conditions of 14.696 PSIA and 60oF,and typically reported in millions of standard cubic feet per day(MMSCFD).

The volume of gas that a screw compressor delivers to thedischarge line is equal to the displacement of the rotors less the"slip" quantity. The term volumetric efficiency is developed to helpdefine the amount of slippage occurring inside the compressor.The following expression is used to define the actual deliveredcapacity of a screw compressor.

ACFM - O, VE

100 (Eq. 3.4)

where:ACFM = actual cubic feet per minutea = Gornpressor displacement in cubic feet per minuteVE = volumetric efficiency in percent

34 ET€T'ffiHHTIEf rl

3.4 Compressor Displacement

The displacement of a screw compressor is equal to the volume atsuction per thread times the number of lobes on the male rotor. Thefollowing expression may be used to determine the compressordisplacement.

o- D'rf!)xcRxRPMC \.D/

(Eq. 3.s)

where:a = cornpressor displacement (CFM)D = male rotor diameter (ft)L = rotor length (ft)GR = geol ratio between compressor and engineRPM = engine speedC = rotor profile constant

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3.5 Volumetric Efficiency

Volumetric Efficiency is simply the ratio of actual delivered volumeby the compressor to the compressor's actual displacement. Sincethere are always losses in a compression cycle, the volumetricefficiency is always less than one. There are many factors thatcontribute to the variance in volumetric efficiency which include butare not limited to:

a. Gas characteristics - specific heat, molecular weight, etc.b. Physica! characteristics of machine - rotor sizes, tolerances, etc.c. Compression ratiod. Rotor tip speede. Slide valve position

To represent the volumetric efficiency of a screw compressor, it involvesthe use of the following relationship:

VE=100- Cr+ SL

[2")xRolz" ) (Eq. 3.6)

where:VECrSLz,zdRo

= volumetric efficiency (%)

= charging resistance (due to temps.) at suction- slip leakage (due to rotor geometry)= cotrlpressibility at suction conditions= Gornpressibility at discharge conditions= corrPrgssion ratio= (k-1)/k (where k=specific heat ratio)

36 EI€-TWffiX&-iTIEII-f

Figure 3.2 is a representation of the volumetric efficiency in a screwcompressor when it is not possible to calculate it explicitly.

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=kt1o->

Pressure Hatio

Figure 3.2 - Volumetric Efficiency for an oil-floodedscrew comPressor

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3.6 Capacity Control

The control of gas flow through a screw compressor is veryimportant both in a refrigeration application to control suctiontemperature as evaporator load varies, as well as in thecompression of natural gas. ln order to control the flow of gasthrough a screw compressor, three common methods of capacitycontrol can be employed. These methods include the use of asuction control valve, reducing the speed of the engine and/or theuse of a slide valve.

3.6.1 Suction Control Valve

The use of a suction control valve is one of the easiest ways tocontrol the capacity moving through a screw compressor. Thevalve works by reducing the absolute inlet pressure into thecompressor, thereby reducing volume. lt is important to know thatthe mass flow into a compressor is directly related to the absolutesuction pressure. lf you can control suction pressure, you cancontrol volume.

There are a few disadvantages to controlling capacity with a suctioncontrolvalve. By decreasing suction pressure, the compressionratio of the machine increases, causing a decrease in overallefficiency. This results in higher wear rates on compressor parts.

3.6.2 Variable Speed

Due to the positive displacement characteristics of a rotary screwcompressor, the capacity of gas through this type of compressorcan be controlled simply through the regulation and variation ofengine speed. With a typical gas driver running these compressors,it is possible to slow the compressor down by as much as 50% withconstant torque input.

The drawback to this type of control is that the operator is unable tounload horsepower if the machine becomes overloaded. Further tothis, the horsepower decline with speed is not a linear relationshipas is the capacity decline. Therefore this type of control is onlyapplicable where excess horsepower is available on the skid.

38 E€TffiEHTIEIT-I

3.6.3 SIide Valve

Often times a great deal of capacity control is necessary. The useof a suction contro! valve and/or the throttling of an engine or motormay not be sufficient enough to cut the required capacity through ascrew compressor. The slide valve now makes it possible to controlcompressor capacity from 1OO% down to 1Oo/o. The slide valveessentially acts as an internal bypass for the gas that enters thecompressor. Figure 3.3 illustrates the flow of gas through thescrew compressor when the slide valve is open therefore capacityis less than 1OO%.

Figure 3.3 - Slide Valve Control

The slide valve does not control capacity in a linear fashion. As the slidevalve is opened, the volume of gas being sent back to suction is a non-linear function of the slide position. For example, if the slide were set to85% load (opened 15o/o), the volume of gas actually being compressed maybe92% of total potential (only 8% is re-circulated). An accurate loadingcurve is important in determining the correct slide valve position in orderto maximize compressor capacity and engine horsepower.

The control of the slide valve in a screw compressor can be done by oneof three ways. The most common method of actuating the slide valve orloading and unloading a screw compressor is through the use of hydraulicpressure differentials on either side of a diaphragm. Other methods ofslide valve control include mechanical and electronic systems orcombinations of these different technologies.

39 E€T€EH-TIEIN

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3.7 Compressor Horsepower

The horsepower requirement for a screw compressor in differentapplications can vary widely. lt is important to be able to accuratelydetermine the horsepower requirements for a screw compressorfor two reasons. First, it aids in sizing an appropriate driver duringengine or motor selection. Second, it ensures that existing assetsare not being over-utilized or under-utilized as field conditionschange in a gas application.

3.7 .1 Theoretical Horsepower

Theoretical horsepower is the power required to adiabaticallycompress a given volume of gas across a pressure differentia!. Therelationship that defines this is seen in Equation 3.7.

where:THP = Theoretical horsepowerP1 = Suction pressure (PSIA)a = Flow rate of gas (MMSCFD)k = Specific heat ratio of gasR = Compression ratio

40 E€'T-&ffiHTIE:If-T

3.7.2 Brake Horsepower

Brake horsepower is the real power necessary to move a volume ofgas through a compression ratio. This calculation takes intoaccount the efficiencies of compression as well as the losses due tomechanical and hydrodynamic friction. This is the true driverrequirement (excluding parasitic loads such as coolers, generatorsetc.). The relationship for determining brake horsepower is given inEquation 3.8.

BHP -rH p*( !) ,[r)IE, ,/ [E, ,l (Eq. 3.8)

where:BHP = Brake horsepowerTHP = Theoretical horsepowerEu = Adiabatic (compression) efficiencyE. = Mechanical efficiency

The adiabatic or compression efficiency of an oil-flooded screwcompressor typically ranges from 35% to 85yo, depending on how the' compressor is configured for the current field conditions. Likewise, themechanical efficiency can vary between 88% and 92o/o, due in part to thehydrodynamic effects of the oil and the extensive lubrication system foundin a screw compressor.

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3.7 .3 Horsepower Calculation Example

Gas Flow

Gas Density

Mol Wt. Of Gas

k-Value of Gas

Elevation

Barometric Pressure

Suction Pressure

Discharge Pressure

Compression Ratio

= 2 MMSCFD@ 14.4psia and 6OoF

= O.65 (compared to air)

= .65 x 28.97(mol wt. of air) = 18.83

= 1.28 (from Figure 1.2)

= 23OO ft

= 13.4psi (from Figure 1.1)

= 30 psig

= 20O psig

= 4.91J =2 {(200+13.a)/(30+13.4)}

**Assuming an Overall Efficiency of 7 SYo **

(zoo+ 1 3.4)

)(#)(so+ 8.4)BHP=

BHP = 22O.5 total compressaon horsepower

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0.75

7

3.8 Adiabatic Efficiency

Adiabatic efficiency is defined as the efficiency of compressioninside a screw compressor for a given set of field conditions andmachine configurations. The determination of this efficiency isdependant on a number of different variables. These include gascharacteristics, temperatures, compressor speed, rotor sizes andmost importantly, the compressor Vi also known as the built-involumetric ratio. Vi will be more thoroughly discussed in thefollowing chapter. Figure 3.4 illustrates clearly the wide range ofefficiencies that are possible under different loading conditions.With the advent of variable Vi machines, the efficiencies of thescrew compressor can easily be optimized over a wide range ofcompression ratios.

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Figure 3.4 - Adiabatic Efficiency of a screw compressor

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4.1 Volumetric Ratio

The screw compressor can be considered to be a volume reductionmachine. Therefore, the volumetric ratio (Vi) is defined as the ratioof gas volumes at suction to the volumes of gas at discharge. Thisratio is a "built-in" feature of the screw compressor, with more ofthe newer machines now allowing the operator to change the built-in ratio in order to more precisely accommodate a wide range offield conditions. The expression for volumetric ratio is seen inEquation 4.1.

Vi -v,

% (Eq 4.1)

where:ViV,vd

= Volumetric ratio= Volume of gas at suction= Volume of same quantity of gas at discharge

By monitoring the field conditions of a screw compressor, anoperator can make periodic adjustments to the Vi setting on thecompressor in order to maximize compression and volumetricefficiencies. Since the parameters measured on the compressorare typically suction and discharge pressures and not volumes, therelationship to calculate Vi is simply the following:

1

_(%)r- l,.P,,Gq a.2)

where:ViPd

P,k

= Volumetric ratio= Discharge pressure (absolute unitsl= Suction pressure (absolute units)= Specific heat ratio of gas being compressed

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)

It is very important to understand that the volumetric ratio is notequal to the compression ratio across the machine, and very Iargeerrors can occur in the setup of the compressor if the Vi is notcalculated properly. The following example illustrates clearly, theimportance of proper Vi determination.

Example Vi Calculations:

o Assume atm. press. = 13.5psia and k=1.25

CASE 1

Gauge Pressure:

P"-175psig =11.7P' 1Spsig

CASE 2Absolute Pressure:

fu_ 188.5psra = 6.61

P' 28.5psia

CASE 3Absolute Pressurewith k-value:

1

[*)'=[

188.Spsra1

)*- 4.53

28.5psia

The following chart should be used on variable Vi machines, inorder for an operator to quickly determine the correct setting for aspecific application. Because the Vi is dependant on elevation andthe properties of the gas being compressed, a Vi chart should bedeveloped for each compressor.

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A unique feature of the screw compressor is that it does notdetermine the finalpressure to which the gas is compressed. Thefinal pressure is actually determined by the pressure in thedischarge line, which is regulated by either a backpressure-regulating valve and/or by the suction pressure of a downstreamreciprocating compressor. The screw compressor does not haveany valves to regulate the discharge pressure of gas during thecompression cycle. The Vi controls the amount of compression thatwill occur for a given volume of gas. As a result, Vi determines thehorsepower requirements and internal pressures that will begenerated during compression. Because the Vi setting on themachine determines how much compression will take place, anincorrect setting will either cause the compressor to over-compressor to under-compress the gas internally. Both of these conditionscause inefficiencies and premature wear in the equipment.

4.1.1 Over Compression

lf the interna! volume ratio, or Vi, is set too high for the prevailingfield conditions, then the gas will remain trapped inside the rotorsfor too long, causing the internal pressure inside the compressor toexceed the discharge line pressure which is set by thebackpressure valve. This is termed over compression, and can beseen graphically in Figure 4.1. ln thiscase, the gas is compressed beyond thesystem discharge pressure, and whenthe gas does finally reach the dischargeport, the high-pressure gas inside thecompressor expands instantly down tothe lower discharge line pressure. Theresult of this scenario is the need formore horsepower to move the samevolume of gas, creating wasted energy inthe form of heat as well as unnecessarilyhigh internal loads on parts such as the psthrust bearings and rotor faces. Overcompression of more than 1ZYois notrecommended by most screwcompressor manufacturers.

Figure 4.1

Suction PortClosed

48

DischargePort Opened

Pd

E€-r*-ffi H{!-{'rtEtr-r

\

4.1 .2 Under Compression

lf the Vi is set too low for the current conditions, the gas inside therotors will not be compressed to a high enough pressure beforereaching the discharge port. This can be seen in Figure 4.2. In thecase of under compression, the gas in the rotors reaches thedischarge port too early and the high pressure gas in the dischargeline flows back into thecompressor. This is knownas back-flow compression.This back-flowcompression causes thegas inside the compressorto reach line pressureinstantly, but at theexpense of higher forces onthe thread at the dischargeport opening and a loss inefficiency moving the gasinto the discharge line.Pulsations and compressorvibration can also be aresult of undercompression.

Suction PortClosed

DischargePort Opened

Pd

Vl = Low

Figure 4.2

ln both cases mentioned above, the screw compressor wi!! stilloperate and move the same volume of gas regardless of overcompression and under compression scenarios. However, thehorsepower requirements will be higher and the wear rates on themoving parts will also be higher, than if the Vi was set properly. Theproper Vi setting will ensure the optimum location of the dischargeport, and maximize the efficiency and life expectancy of thecompressor.

49 E€'}*ffiHHTIEIFI

4.2Va and Slide Valve Movement

As was discussed earlier, hydraulic actuators typically control themovement of the slide valve. The Vi is adjusted by a manual threadadjustment on the suction side of the compressor. When thecompressor is fully loaded, the slide valve is pressed up tightagainst the Vi stop. When the slide valve and the Vi move axiallytogether, the compressor's Vi is changed. Figure 4.3 illustrateshow these two components work together.

Higher Vi fl

Lower Vi

Figure 4.3

50 EI€T*-ffi T=[-"{TIEIN

SIide

Slide

Alternatively, when the compressor needs to be unloaded, the slidevalve moves independently of the Vi. When a screw compressor isunloaded through the use of the slide valve, the Vi no longerfunctions properly. When the slide valve is opened up, thecompressor's built in volume ratio becomes a constant andemulates that of a "lo\ r" setting. ln a high compression ratiosituation, it would be very inefficient to run with the slide valve openfor extended periods of time, due to the high degree of undercompression that would occur. See Figure 4.4 for an illustration ofthis.

Fully Louded

Unloaded

Figure 4.4

51 E€-T-EEHTIEI-T

Rotor

Slide

Slide

52 EI€ET.E5 EHTTEIT'I

5.0 OIL MANAGEMENT

v

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5.1 lmportance of Oil

Oil plays a vital role in the oil-flooded screw compressor. Thecontinuous injection of oil during machine operation plays threemajor roles. Firstly, the oil cools the heat due to compression, andallows the control of discharge temperature. Secondly, the oillubricates al! of the rotating parts, ie. bearings, gears etc. Finally,the oil creates a hydrodynamic seal between the male and femalelobes in order to prevent metal-to-metal contact of the spinningrotors and minimize blow-by of gas inside the case. All of thesefunctions of oil are necessary for the compressor to performproperly. The type of oil used, either synthetic or mineral, is highlydependant on the service and composition of gas that the machineis compressing.

Synthetic oils (Polyglycol) are used mainly in dry gas applications.These oils have the ability to resist hydrocarbon breakdown. lf thedischarge temperature remains high, these oils are also acceptablefor use in a wet gas environment.

Mineral oils are used in wet gas applications. These oils arecheaper to use, but have an affinity to absorb the gas. lf a mineraloil is used inside the compressor, a careful oil maintenance planmust be in place to monitor the dilution of the oil over time. With thedilution of oil, comes viscosity breakdown, leading to highertemperatures and increased wear on the compressor.

The oil system on board a screw compressor is a closed system.This means that oil should not have to be added to the reservoirvery often. The oil is continuously recycled from discharge, througha cooler and back to suction. lf the oil level is constantly getting low,then the oil is either moving down the sell line indicating a collapsedfilter in the oil separator, or there is an oil leak somewhere else inthe system.

Oil is injected at various locations within the compressor. Table 5.1identifies the importance of oil throughout different parts of thecompressor. Oil is injected at the suction port; it is also injected viathe slide valve, as well as through the bearings into the rotor area.The typical injection flow rate for oil in a screw compressor isapproximately 7 - 9gpm of oil per 100CFM of capacity and is alsodependent on the differential gas pressure. Too much oi! beinginjected can cause an increase in horsepower consumption. andnot enough oil injection can cause premature compressor failuredue to lack of lubrication and high discharge temperatures.

53 E€A*ffi#E**:-rrEtr-r

JournalLubrication

RadialBearings(Side & Main)

Oil wedgeHeat removal

ThrustBearings

Thrust force offset

BalancePiston

BalancePiston

Thrust force offset

Oil lnjection Rotors Rotor sealCompression heatsuppression

MechanicalSealLubrication

MechanicalSeal

Oil wedgeHeat removal

GearLubrication

lntegral Gearwhereapplicable

Gear lubricationHeat removal

CapacityControl

Slide valve Hydraulic capacity control

Table 5.1 - Lubrication and the Effects

54 E€T€ffiHTIEfT

5.2 Temperature

A temperature rise in a screw compressor is of significantimportance for the successful operation and maintenance of thecompressor package. During the compression cycle, heat isgenerated continuously along the rotor lengths. The temperaturerise in an oil-flooded screw compressor is controlled by theinjection of oil throughout the compressor housing. lf temperaturesinside the compressor become excessive, problems with oilbreakdown, deposits and thermal stresses may develop, leading tothe premature failure of the unit. Furthermore, if too much oil isinjected, or the injection temperature of the oil is too low, problemsof condensation inside the equipment may occur. lt is veryimportant that an adequate amount of oil is injected at the righttemperature in order to maintain an appropriate dischargetemperature.

There are severa! factors that must be addressed when determiningthe temperature rise of the gas inside the compressor. First, andmost important, the type of oil used and its physical properties (ie.specific heat, density, etc.) play a vital role in the amount of heatthat it wil! remove during compression. Along with the type of oilused, the amount of oil injected will also have an effect on the endtemperature. Molecular weight of the gas, compressor RPM andcompression ratio a!! have a hand in the determination of thedischarge temperature; which in turn determines the oil systemcharacteristics. The lighter the gas, the higher the temperature risewill be for a given compression ratio, as wel! as the higher thecompression ratio the higher the temperature rise. There are manyfactors that influence the temperature rise in a screw compressor.Equation 5.1 can help estimate the rise in temperature of the gasinside the screw compressor with oil injection.

55 E €fr*-ffi ffiF-{ Tt Er-r

-T1 +[t+)'"'

-']xrxE,

E"(Eq. 5.1)

where:Tr = Suction temperature ('R)Tz = Discharge temperature ('R)Pr = Suction pressure (PSIA)Pz = Discharge pressure (PSIA)E. = Cooling effects of oil (approx 50%)Eu = Adiabatic efficiency (approx 75%)

5.3 Keys to Long Compressor Life

5.3.1 Oil Management

Good oil management is essential in the proper operation of thescrew compressor. Proper oil management includes keeping thetemperature of the oil steady in order to maintain proper viscosity.As well, it is imperative that the oil pressure in the system remainsat some point above gas discharge pressure in order for the oiltoflow through the machine. Finally, the quality of the oil should beoccasiona!!y monitored both visually and through a good oilanalysis. This is to determine if there are any internal problems thatmay begin to appear such as particulates in the oil, high dilution ofoil, etc.

5.3.2 Proper Vi Settings

Maintaining the correct Vi is critical for the compressor to achievehigh levels of run time without failure. lmproper Vi settings willresult in higher than necessary wear rates as well as increasedpulsations and vibrations in the equipment. Periodic checks of theVi should be made, especially as field pressures change.

T2

56 El-E'-T*-ffifit EHTII=n

5.3.3 Shut Down Procedures

lf the compressor is shut down for more than three to four months aspecial procedure must be followed to ensure a safe and successfulstartup in the future. First, shut both the inlet and discharge stopvalves. Then disconnect the power supply and al! air sources.!nside the contro! panel, place a moisture-absorbing compoundsuch as silica gel to preserve the panel. lf there is water-cooling onboard, drain allwater and other liquids from the system ie. from theinlet separator, to prevent corrosion. Place warning tags on thecompressor control system, starter and all the closed stop valves.Finally, every month the oil pump should be run (where applicable)and the compressor should be turned by hand ten to fifteenrevolutions. Prior to the startup the oil must be changed for a freshcharge into the system.

57 E€-tr*€EHTIEI-I

6.0 MAINTENANCE AND TROUBLE.SHOOTING

58 EI.EETEEHTIEII'I

6.1 Maintenance

The implementation of a good maintenance program wil! ensure thatyou will be able to maximize the compressor run-time as well asminimize the maintenance dollars spent on the asset. Table 6.1 is aguideline for the maintenance and operation of a screwcompressor.

Table 6.1 - Operation and Maintenance Schedule Guidelines

Operation Time lnterval

59 E€'T'-ffi EHTIEII-I

Suction Pressure Daily

Discharge Pressure Daily

Suction Temperature Daily

Discharge Temperature Daily

Lube Temperature Daily

Lube Pressure Daily

Lube Differential Pressure Daily

lnlet Differential Pressure Daily

Lube Level Daily

Engine Speed Daily

Low lnlet Pressure Shutdown Monthly

Low Discharge Press. Shutdown Monthly

High Discharge Press. Shutdown Monthly

High Oil Diff. Press. Shutdown Monthly

High Discharge Temp. Shutdown Monthly

High Oil Temperature Shutdown Monthly

Low Cooling Water Press.Shutdown

Monthly

Sample Lube Oil to CheckAppearance and Run an OilAnalvsis

Every 1,000 hours for the first6,000 hours and then every2,OO0 hours thereafter

Change Lubricant Every 2,000 hours or 3 monthsup to a maximum of 6 months inconiunction with an oil analvsis.

Change Oil Filter Whenever oil is changed ordifferential pressure is greaterthan 10psi

Clean Oil Strainers Whenever lubricant is changed

Check Noise Level Daily

Check Slide Valve Actuator andSettings

Monthly

Check Vi Setting (variable Vi unitsonly)

Monthly

Check Electric Motor Bearings Yearly

60 E€'T*.$ffiEHTIEN

Lubricate Electric Motor Bearings Yearly

Check Coupling Alignment Yearly

Tighten Mounting Bolts Yearly

lnspect Rotor End Play Every 6 Months

lnspect Oi! Separator Mesh PadElement

Every 6 Months

lnspect Oil Separator FilterElements

Every 6 Months

lnspect Cleanliness Every 3 Months until requiredcleaning frequency isestablished.

Replacement of Seals andBearings (radial and thrust)

Every 3 years or 24,OOO hours

61 E€-I*-ffi[ EH -rlEtt-t

6.2 Troubleshooting

Troubleshooting a compressor can be one of the most frustratingexperiences to work through due to the fact that there are so manydifferent variables that can contribute to an equipment problem.The following chart is presented as a guideline to work with introubleshooting a screw compressor. This information has beencompiled from different manufacturers and field service reports,and is not to be used as the solution to all problems with a piece ofequipment. As with any problem, gather all relevant informationand make visual and audible inspections in order to prevent theunnecessary replacement of expensive compressor components.Some of the more obvious things to look and listen for are:

a. Damaged tubingb. Loose wiringc. Any loose fittingsd. Unfamiliar compressor, pump, or engine noises

Table 6.2 - Troubleshooting

COMPRESSOR SYMPTOM PROBABLE CAUSE

62 E€T*€trEHTIET-f

Compressor fails to start 1

23

456

A protective switch is tripped.Class B timer is activated.No power or air supply tocontrol circuit.Bad or wrong connections.Bad or blown fuses.Defective pre-lube pump(where applicable).

Compressor shut downimmediately after starting

1. Low oil pressure.2. Cold oil.3. High discharge temperature.4. Shut down switch(s)

malfunctions.5. Oil filter differentialtoo hiqh.

Compressor does not load orunload

1. Leak in control lines.2. Restriction in control Iines.3. lmproperlv adiusted control.

Low oil pressure 1. Plugged oil strainer2. Plugged oil filter.3. Low oil charge.4. Low oilviscosity.5. Worn oil pump (where

applicable).

High oil pressure 1. Cold oil.2. Oil pressure regulator is not

set DroDerlv.Low oil temperature 1. Thermal valve element is

defective.2. Oil heater or thermostat

defective.High oil temperature 1. Thermal valve element

defective.2. Cooler fan not working

(defective start switch).3. Dirty oil cooler.4. Low oilviscositv

H igh d ischarge temperature 1. High injection oiltemperature.

2. Plugged oil strainer.3. Abnormal operating condition

(high discharge pressure orlow suction pressure).

4. Wrong Vi.5. Low compressor oil.6. Mixing valve defective.7. Plugged oi! filter.8. Dirtv oil cooler.

Low inlet pressure 1. Excessive pressure drops inlines.

2. Capacity control notmodulatino.

High oi! consumption 1. Oil not returning tocompressor.

2. Defective or improperlyinstalled oil separatorgaskets.

3. Excessive oil charge insvstem.

63 E€-E*€MHTIEN

Motor runs hot 1.

2.3.

4.5.

6.

Too many starts within ashort period of time.Excessive current draw.Low voltage (never less than9OYo of nameplate rating).Restricted ventilation.High ambient temperature(>1 1O"F).lnsufficient or excessive lubein bearings.

7. Defective bearings.

Excessive compressor vibration 1. Wrong Vi.2. Loose anchoring.3. Misalignment.4. Oil compression (excessive

lubrication).

64 E€,*-.@HHTIEI-I

7.O PERFORMANCE CURVES AND EQUIPMENTMONITORING

65 EI€TT€'EHTIEII'I

.-,

7.1 Loading Curves

The following graph represents a typical loading curve for a screwcompressor in natural gas service. The curve shows the maximumcapacity of the compressor along with the horsepower utilized atvarious suction pressures. The curves are typically developed withmaximum engine RPM and a constant discharge pressure.

66 ET€'"f*€:HTTEIN

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67 EI.ET€EHTIEN

7.2 Compressor Evaluation Sheets and Diagnostic Reports

Attached are the latest evaluation forms for the screwcompressor. Page 69 contains a copy of the field data sheet thatoperators will fill out for analysis. When filling it out, it is veryimportant that the data is collected with good accurate gauges.As well, gauges should be read to the degree of precision thatthey were designed, in order to prevent errors in the reports.Page 70 is a copy of the report that is generated based on theinformation provided in the field data sheet.

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P.ET...TY E EITIFR.ESSER, I"III.H.A.EEi.IEH T

Company tlanadian l.latural Resourees LimitedI'lodule Humber 90f4Compressor Hame CnLLIHG LnHE l-11Eompre=r'rr Lo,:ation LSn 01-1'l-7rl-t3 W4Receigt Point Location HOVL STATIEH 1d4?

Eleuation -fa=l l0fE.1

Frame l,lanufacturer

Franre Llodel

T]treProfi I e

Eoter Iiameter[mmJL/tr RatioGrp Size [in.]tiear Eatio

Mycom

16[VLREHel i ctsl

asymmetric1E3.t

1.65

[.[032.te

Thermal Trans.35llll Series

9074IEffiiic trJ-06;leoI E"*-o*I ,lrI PrnductionI rr

I r:uerirrur

I oo*rrir*

ilantr{acturer[',lodel DescriptionI'laximum EPl,l

I'liri mum Horsepo,r+erRunning 5pe*dHorsepo+er Used

Co,tpressor

Eotor EiameterRotor Lengthtiap Size

Temperature SuctionTemperature 0ischargeFressurE 5usti,f,nFressure EischargeCompression Ralio0ptimum Ul SettingPort 9etting or Fired tiSlide 1lalue Prsition

c Effi+iencyMechanical EfficiencyCompression Efficien+yDuerall Effi+iencyThrust Load0uer/Under CompressionCompre=sor EPI'l

Compressor Flow

0uer EompressienUnder CompressionHigh Tempereture GasSuction Prpssure EeclinaCoalesing FilierSlide lblue PositionHigh Thrust LoadBeering Fatigue LifeDste of La5t Ga= ,{nalluis

Horaepo+er [JtiliredEoosttlapacity LJtilieedlncremerilal ProductionElpass Dpen

R6.Tlf,Ll . DrtEiolTtE l{tHtRE VEE!t{i 1.!

Czterpi I I arESSOETA

I S00

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180

Llanr:facturerI'lndel HumberHorsepou,,er ErawLlax Design PressureLlix Eesign Temp

Comp''essor Oil

0il Fressure0il Filter Oifferential

[il lnjection Temp

0il Erit TempEalance Piston Pressure

Bil FlovHeat Rejected0il L,lrcosity

Last 0il Filter Ehange

Fuel Gas LHVEas k-trtblue

st. Fuel fonsumption

Or+rhruJ HoursComp 0uerhaulEnEine TtrF End

Engine Eottom EndT,rtal lime sn Unit

Corrrpressor 0il0il Pressure

High Oil lnjection TempHigh Balance PiEton PressureLrv Ealance PiEton Pressure

Itlin ltlachine Oiff '+rJo Pump[,,lin l,lachine Biff v/ Pump

tlil Filter Eifferential

5 HiUHLTSIS REOUIREO"

'Eeconrnended Hi

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e.e

70 E€-Tr.ffiSXHTIEI-I

71 EI€IT.E EHTIEIT'I

8.0 HANDY CONVERSION FACTORS

t*

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8.1 Handy Conversion Factors

The following conversions may come in handy when workingbetween metric and imperial systems of units on a compressor.

PRESSURES

1 psi = 6.8948kPa

1 mmscfd @14.696psia & 6O"F = 28.26E3M3 101.325kPa & 15"C

TEMPERATURES

To convert from Centigrade to Fahrenheit:

"F = t(9/5)x("C))+32

To convert from Fahrenheit to Centigrade:

"C = ("F-32)x(5/9)

72 EI€T-EEM'T.!ETI-T