98-GT-32 - ASME Digital Collection

THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS345 E. 47th St, New York, N.Y. 10017 98-GT-32

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Copyright OD 1998 by ASME

All Rights Reserved

Primed in U.S.A.

THE CENTRIFUGAL COMPRESSOR INDUCER

C Rodgers

3010 N Arroyo Dr,

San Diego , Ca 92103

ABSTRACT.

The function of the centrifugal compressorinducer is to provide wide flow margins fromthe design flow to the stall and choke flowlimits, together with high static pressurerecovery. At transonic conditions the inducerbecomes the critical to impeller performancein that shock losses and blockage growthdiminish stable flow range and may triggernear vertical stage characteristics.The paper covers the various types of inletconfigurations upstream of the inducer,followed by a review of the state-of-artinducer design for centrifugal compressors,culminating with some research developmentsin transonic inducer blading characteristics.

NOMENCLATURE

A AreaAt Inducer throat areaC VelocityCp Specific heatCFS Volume FlowD DiameterETAC Compressor Efficiencyg GravityH Headk Ratio of specific heatsL/E Leading EdgeM Mach numberMu Delaval Number = U2 /(gIcRT)"N Rotational speedNs Specific speed

= o I CFS / ( g Had ) 0.75o Throat openingP Stagnation or total pressure

Q

Compressible flow functionW4T/ AP

Qcrit Compressible flow function ,sonicconditions

q

Work factor, or dynamic headRMS Root mean square

Blade pitchT

Total temperatureU

Blade speedW

Massflow or relative velocity

a Air angleBlade angle

A Differencec Diameter ratiori Efficiency

Flow factorp Densityco Angular velocity

Subscripts

1 Inducer inlet2 Impeller exitad Adiabaticc Compressore Stage exith Hubs Shroudt Throat

Note : All angles are with respect to theaxial or radial direction.

Presented at the International Gas Turbine & Aeroengine Congress & ExhibitionStockholm, Sweden — June 2—June 5, 1998

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1. INTRODUCTION.The centrifugal impeller stage characteristicis basically determined by its entry and exitblade shapes. The entry or inducer sectionthroat contols the stall and choke flows, andthe exit or exducer essentially governs themomentum exchange.The demand for more efficient compressoroperation at higher Mach numbers andextended suction specific speeds of pumpsduring the early 1950's, focused extensiveresearch on separate axial inducers ahead ofthe radial section of the radial impeller. Inparticular these inducers were shown toprovide improved centrifugal compressorefficiencies at the higher pressure ratiosnecessitated by gas turbine cycle performancerequirements.The past restrictions to successfully analyzethe impeller internal flow promptedmodeling the inducer section as its axialcompressor stage counterpart, a practicewhich is still current, for at least meanlinestage performance predictions.Later performance improvements incentrifugal compressor technology emergingin the last two decades have been realized by:

. Adoption of diffusion controlled three-dimensional backswept impellers.. Applied transonic axial compressortechnology.. Numerically controlled machining for three-dimensional arbitrary blade surfaces.. Refinement of meanline performanceprediction codes together with application ofcomputational fluid dynamic codes.This paper covers the various types of inletconfigurations upstream of the inducerfollowed by a review of the state-of artinducer design, culminating with recentresearch developments in transonic inducerblading characteristics.

2. INLET CONDITIONS.

Maximum performance of the inducer isachieved with unspoiled inlet flow.Centrifugal compressor inlets are typically ofeither the axial or radial single entry type. Insome instances the installation may requiredual, or bifurcated inlets which may imposeflow distribution problems, and or, additionalpressure losses dependent upon specific

geometric features. Maximum performance isnormally obtained with an axial bellmouth typeinlet. In some cases the radial inlet withcircumferentially accelerating flow around a 90deg bend towards the inducer eye is oftenselected, such as gas turbine engine andmultistage centrifugal compressorconfigurations. This is not to be confused withthe so called "radial inlet" from a suction pipe.Several types of centrifugal compressor inlet

configurations are shown on Figla through f.Ref 1 describes the results of tests concerningradial inlet geometric features, and inparticular the influence of axial lengthextension upstream of the inducer leadingedge. The test results shown on Fig 2 indicatesthat relatively short axial extensions can beused at De Laval Mach numbers up to 1.15,without significant performance penaltiesproviding;

. The ratio of the hub axial extension to thehub diameter L/ D1 5 >0.4. The mean streamline velocity maintains apositive acceleration gradient, especially inthe last third of the flow path up to theinducer leading edge.. A meanline axial velocity ratio of 2.0(exit/entry) is maintained.

The use of these simplistic guidelines forpreliminary design layouts provides a quickdesign aid, but it is not recommended as adetail design tool, especially when inletdistortion may be present at the radial entry,and or, higher inducer tip relative Machnumbers are incurred. Under such conditionshigher meanline acceleration rates arepreferred to prevent excessive shroudboundary layer growth, and a complete inletthroughflow analysis is warranted.The effects of inlet flow distortion oncentrifugal compressor performance areaddressed in Ref 2. Radial flow distortion wasfound to have a more pronounced effect onsurge margin and efficiency thancircumferential distortion. It is generallyaccepted, however, that the centrifugalcompressor is less sensitive to flow distortionthat its axial flowcounterpart.In most compressor applications it is requiredto provide an inlet plenum or muff to directthe airflow from an exterior source to thecompressor eye, the radial type of inlet is

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often used in conjunction with a rectangularinlet muff.The area ratio between the muff inlet andcompressor eye is again a primary factor inmuff sizing, and performance both in terms ofpressure loss and flow distortion.Area ratios of at least 3.0 to 4.0 are requiredfor minimum loss, and can assist in reducingsensitivity to externally propagated flow

distortions.Asymmetric or tangential type suction nozzlescan create non-uniform flows with prewhirlcomponents at the compressor eye. The use of

1 Muff

a flow splitter at bottom dead center iscustomary, especially with confined low arearatio configurations.Although an inlet center body or spinnerahead of the impeller is esthetically pleasingmost turbocharger centrifugal compressorsare produced with an uncovered impellerretainment nut. Comparative compressorperformance tests with and without astreamlined spinner al m\---111-2< 0.4) haveshown negligible differences.

L JFig 1 Centrifugal Compressor Inlet Types

Lid1 Impeller Effy%Ihr 0 62 82.60 51 82.00.46 79.7

I dl

(----- Mu=1 .3

Fig 2. Effect of Inlet Axial Length

Fan Interstage Separate Inducer 3D

Mixed Flow

Fig 3 Impeller Types

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3 REVIEW OF INDUCER STATE-OF-ART.

The main function of the impeller is toprovide the required pressure and flow levelswith minimum energy input and with stableflow operation. The desired pressure level isidealistically attained merely by providingsufficient tip speed, with the product ofrotational speed and tip diameter, whereasthe required flow and stability is mostlydependent upon the correct shaping andsizing of the inducer section.The inducer is the critical section of the

impeller, especially for efficient operation athigh inlet relative Mach numbers.Essentially five impeller entry (inducer)configurations have evolved during the courseof centrifugal compressor and pumpdevelopment as depicted on Fig 3.The simple radial and diagonal leading edgesare still utilized on low pressure fans and forlow specific speed applications, exhibitingcomparable performances to inducer type

impellers at Mum40. 6.The demand for more efficient compressor

operation at higherMach numbers andextended suction specific speeds of pumpsduring the late 1940's, focused extensiveresearch on separate axial inducers ahead ofthe radial section of the impeller.The effects of inducer axial chord, splitters,camberline shape (circular and parabolic),leading edge contour was some of the manyinducer geometric features studied.Analysis of these experimental studiesrevealed the similarity of inducer operatingcharacteristics to those of the axialcompressor rotor stage, even though followedby an impeller flow turns from the axial tothe radial direction. It was subsequentlypossible to correlate inducer characteristicswith extensive existing cascade data of axialcompressors as discussed in Ref 3.Initial meridional sizing of the inducer isdependent upon;

.Inducer hub diameter Dm as constrained byeither blade manufacture, or shaft diameter.inducer tip diameter Di s for minimumrelative curvature concerns.For zero prewhirl the inducer tip relativevelocity is given by;

Wis = m 2 +[ Massflow pi Alf) 0.5‘-lsExamination of the above equation shows thatfor constant massflow and density that theright hand term increases with increasingdiameter, whilst the left hand term decreaseswith increasing diameter.The relative velocity Wis therefore has aminimum which for incompressible flow is ata tip relative flow angle f3 is of 55 deg. Forcompressible flow, Wis is minimum with [3 i sbetween 55 to 65 deg, depending upon theinducer hub to tip ratio, and specific speed.High specific speed designs may howeverresult in excessive shroud curvature (1/- 15—>Dib), thus selection of a lower inducer tipflow angle (than optimum) may beconsidered.Having made a preliminary selection of theinducer hub and tip diameters attention isfocused towards sizing the inducer throat topass the desired flow, of sufficient area toavoid choking , or cavitation in pumps. Achoke flow margin of up to10% is typical forMu > 1.5 compressors, the appropriate throatof which, must be compatible with a selectedblade entry angle and possess adequate stallmargin. The following conditions shouldtherefore be satisfied;

. Smallest possible inducer hub diameterconsistent with blade manufacture and shaftdynamic concerns.. Optimum inducer tip diameter for minimumtip relative Mach number and tolerableshroud curvature.. Computation of a corresponding hub, RMS,and tip velocity triangles. An inducer choke margin of up to 10%, withcontrolled blade turning between the leadingedge and the throat.. Choice of a radial distribution of bladeincidence giving adequate stall margin.

Two axial compressor blade designprocedures adopted in the design of inducersfor centrifugal compressors were thecorrelation maximum inlet relative Machnumber versus blade throat opening to pitchratio, and blade stalling incidence as afunction of turning and solidity (Ref 3).

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Typical inducer choke flow data is shown onFig4, and parallels the theoretical choke lineup to near the sonic limit, after whichmaximum flows are highly dependent uponsuch factors as blade thickness, leading edgeradius, and blade turning between the leadingedge and throat, as will be discussed later.The iteration of the throat opening and theblade inlet angle to provide adequate chokeand stall margins, with as is often, the desireto have blade elements stacked radially isresolved computationally. As a designguideline the one-dimensional stallingincidences (unblocked annulus) shown on Fig5 (Ref 4), may be used.

ca

v

4tp ifrckness

from&

ng types

RelMach noM 1w

Theo etical

Test

Thro t Op • c.--

to Pitch Ratio °is cos 0:1

0.4 0.6 0 8 1 0 1.2

14

Fig 4 Typical Choke Flow Correlation

.4

.2

h Non m iws

Based onAnalysis

1D

8

6

.4Tip Incidence deg

0 5 10 15 21

Fig 5 Typical Inducer Stalling Incidences

Note that for a radial leading edge, radialelement blading ,the blade angle is prescribedby

tan 01 tan I is (D1 D1S )For zero prewhirl ,near uniform entryvelocity, and solid body rotation, the radialvariation of incidence with this blade angledistribution becomes minimal.Curtailment of axial length in multistage

centrifugal compressor configurations forshaft stiffness reasons generally restrain theimpeller to configurations shown on Fig3,where the leading edge lies close to anorthogonal. At lower speeds (Mu < 0.5)maximum flows depart from the choke flowcorrelation (Fig 4), negative incidencesincreasing beyond 20 deg. As a general ruletherefore a maximum negative incidence is setin these cases.In many instances alternate inducer bladecutbacks are preferred to maximize thethroat area for a given blade angle and tipdiameter, or simply for ease of manufacture,especially small turbocharger type impellers.Alternate blade cutbacks , or splitteredinducers provide a larger throat area for agiven blade angle and therefore tend toexhibit a wider flow range between impellerstall and choke, sometimes at the expense ofslightly lower impeller efficiency (of the order1% point), as compared to fully bladedinducers.Equal circumferentially spaced splitteredinducers result in unequal channel flows as aresult of the velocity gradient from suction topressure surfaces, and have been postulatedto cause additional mixing losses at theimpeller tip. Some compressors have splittersspaced for equal channel flows and claim tohave improved performance.Tandem cascadetype inducers have also been developed, Ref5, in an attempt to improve compressor flowrange with indifferent results.A potential disadvantage of splitteredinducers is increased noise generation ifreduced blade number drives the bladepassing frequency into the susceptible audiblerange.As mentioned previously, compressoroperation towards shut-off results in impellerflow recirculation back to the inlet causingthe inlet temperature to increase rapidly.

1.4

1.2

1.0

0.8

0.6

0.

1

1Ri

Ma

1

0

0

0

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The flow recirculation first appears atinducer stall, and may be sensed with aThermocouple, suggesting this as an activestall sensor. This inducer stallingcharacteristic has been delayed in many casesby the use of inducer shroud bleed, asdiscussed later.

4. ADVANCED TRANSONIC INDUCERTECHNOLOGY.

The thrust of recent advances in single stagecentrifugal compressor technology has beenmotivated to a large extent by the aerospaceand automobile communities in the interestsof increasing power density through higherairflows per unit frontal area (higher specificspeeds) and improved thermal performancesat higher stage pressure ratios. Single stagepressure ratios of 15.0 with very reasonableefficiency and adequate operating range, havebeen achieved by Canadian Pratt andWhitney. Excellent performance for a small4.0 inch tip diameter impeller at a pressureratio of 6.0 is reported in Ref 6 . In parallelwith aerodynamic achievements, improvedmaterials and structural design techniquesare being used, providing the capability ofoperating at higher tip speeds,with increasingcyclical life.

A tripartite focus of advanced transonicinducer research has recently been:. Adoption of inducer shroud treatments. Transonic Blading controling inducerincidence and diffusion. Incorporation of leading edge sweep

4.1 Inducer Shroud Bleed.The use of inducer shroud bleed adjacent tothe inducer throat has in many instancesshown to be effective in extending compressorflow range by either decreasing surge flow,and or, increasing compressor choke flowalong a constant speed line.Refs 7, 8, and 9, respectively discuss theeffects of shroud bleed on a high pressureratio gas turbine centrifugal compressor, andsmall turbocharger compressors.Circumferential bleed hole ports or slotslocated adjacent to the inducer throat permitflow to be either blown out, or sucked in, asdependent upon the port internal staticpressure,and thus throttle position. In effect

active flow in, or out of the port, appears tobe regulating the effective inducer throatarea, but the flow mechanism is a littlesubtler as will be described later. Thedischarged inducer shroud bleed may beeither simply recirculated back to the inlet,or dumped externally.Externally dumped shroud bleed was used onthe small high pressure ratio compressordescribed in Ref 6, where increasedcompressor work factor with increasingshroud bleed was speculated to stem fromreduced impeller tip flow blockage. Note thatRef 6 tests were conducted with the bleedflow dumped externally, whereas in Ref 8tests the bleed flow was recirculated back tothe inlet, furthermore with a vaneless diffuserinstalled.

Fig 6 Research Impeller

These beneficial effects of impeller shroudbleed instigated the subsequentimplementation of bleed capability on ahigher specific speed experimental researchcompressor partially described in Ref 10, andshown on Fig 6. The overall stageperformance of this research compressor withrecirculated and dumped shroud bleed ascompared to baseline conditions with asmooth shroud having no bleed holes is shownon Figs 7 and 8 respectively.The pronounced effect of shroud bleed oncompressor flow range is readily apparentand is much more significant than the minorchanges in efficiency.Measurement of the dumped bleed permitteda flow and enthalpy balance at the impellertip and assessment of impeller and diffuserperformances in comparison to the baseline

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shown on Figs 9,10,and 11,which respectivelyreveal:. Increased work factor and stability (highernegative slope of work vis flow) of theimpeller with shroud bleed. Increased impeller efficiencies.. Essentially the same diffuser recoveries, yetpermitting operation on the stalled (positiveslope) half of the diffuser characteristic.Ref 11 reports that a continuouscircumferential bleed slot provided improvedcompressor performance as compared tobleed holes, which could be due in part to,reduced discharge coefficient and slightlyhigher bleedflow. It would be anticipatedthat the choke flow increase is proportionalto the bleed port area (of the order 10-15%At), a contraction coefficient multiplierrepresentative of the effective orifice, and thebleed port pressure differential. These doindeed influence the bleed flow but the majoreffect of shroud bleed is summarized asfollows:

Induced Flow. When the inducer shroudstatic pressure becomes sub atmospheric flowis sucked inwards. This has the effect of

decreasing the impeller tip exit blockagewhich increases the work factor, both forrecirculated and dumped bleedconfigurations.

Expelled Flow. With increasing backpressure, towards stall, the shroud staticpressure rises above suction and flow isexpelled. This also has the effect of decreasingthe impeller tip exit blockage, with increasesin both impeller efficiency and work factor.The effect of recirculated and dumped bleedconfigurations is different however, in thatrecirculated bleed mixing with the suctionflow reduces the effective corrected speed andthus the overall pressure ratio. Accordinglythe overall efficiency as obtained from thesuction and discharge pressure andtemperatures generally falls below that of thedumped bleed configuration.The intrinsic process by which shroud bleeddecreases the impeller tip blockage is notidentified, and is not necessarily Machnumber related. It is surmized to be morerelated to the shroud curvature, and thusmore effective on higher specific speedimpellers.

u

701Efficiency

%,/ \ "..1-'°-..\-S. \)C"..-

\

'''\'

\

1------■Test MapCompressor

Research S—— -----,

in_ 6.5 inch lip dia7 ----.. --.,Zero Bleed -

- - - Bleed Recirculated \ 1.64

3 0 ~Pressure 1 1.56

Z.0

Ratio_ t-s

)N, 1.4f

1.31 Mu

0.98 QJQcrit10

04

0.6

0 8 .0

80 Efficiency%

1

\ /----- -5.--- ■•••

70 I

Test Map ResearchCompressor

g . o_ 6.5 inch Tip diaZero Bleed

_ Bleed Dumped '\ 1.64

3 0 1.56Pressure1.4f

Ratiot-s 1.31

2.0 - ---,....,

0.98 QfOcrit10 _ _

04 1.0

Fig 7 Effect of Recirculated Bleed

Fig 8 Effect of Dumped Bleed

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No Bleed - - - Bleed

\0.98

1.310\8 \

1.4\ '1.64

1.56 Mu

0.6

0.8

1.0

1.0

Diffuser

Recovery 0 5Cp 2-e

0Flow Mail 0 . 4

Fig 9 Effect of Bleed on Diffuser Recovery

\

\

\-,-, -.. - _----__c ---- ----.

N\ 0.98

...t-:..„.....„.t.....4...„,....._----.-------

1.31 .48 -1 .1.54

Mu

1.0

/,./r7 --''`:;; --;%:---":---------------Mu

/0.98 1.31 1 8 1.56 1.64

Flow 0/Qcrit .8

1.0

Fig 11 Effect of Bleed on Impeller Efficiency

Work

Factor

q

0.8

0.7

0.6

Flow 0/Ocrit 0.4 0.6 0.8

Fig 10 Effect of Bleed on Work Factor

90

ImpellerEfficiency

% 80

70

4.2 . Transonic Blading.Impeller efficiency is highly dependent uponinducer tip relative Mach number whenMims exceeds unity. As a consequence in thetransonic range with optimum specific speedsbetween 0.6 and 0.9, Micas is often used asthe single major parameter correlation ofpeak impeller efficiency, Fig 12,andaerodynamic blading design becomes focusedtowards minimizing the effect of itsmagnitude.Design of the inducer blading from theleading edge to the throat poses a dichotomyin that in order to maximize range(choke/stall flows) it is necessary to maximizethe throat area at choke ,yet effectivelyminimize the throat area at stall.

Although a solution is possible in subsonicblading with camberline turning from theleading edge to the throat, operation attransonic conditions leads to excessive shocklosses. On the other hand reducing theturning for a given throat area decreases theblade inlet angle, and thereby increasesincidence for a given flow.Comparative inducer blade angledistributions for subsonic and transonicdesigns are shown on Fig 13, indicating thatslight inflexing of the initial turning may bebeneficial in reducing overshock strength.

The results of comparative choke flow testswith subsonic parabolic camberline inducersand transonic reduced entry turning are

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81 2 1 3 1 4 1 5

Fig 12 Effect of Tip Mach Number on Impeller Efficiency

92

90

86

88

8

0

0

0

ImpellerEfficiency

•,o0 o

°00 0o

e

%.t-t o0

o0 0

o r 0

1D UniformVelocity

o Test Data. Several

Inducers

Axial o

o 0

Transom:

I Inducer Tlp Relmive Mach

c

Number

lnflexed Turning Thin Blades1.4

12Rel

Mach n

1.0M1s

1.4

shown on Fig 14, and reveal an intriguingresult in that a transonic inducer can exhibithigher choke flows with a smaller throat areaat the higher Mach numbers. This is a resultof reduced shock losses and blockageaccumulation up to the throat, both fromreduced turning and thinner blading.Thinner blading is however a detriment tocost, durability and vibrational concerns,which may dominate in some applications. Aninteresting solution is the partial thinning ofthe inducer blading from leading edge tothroat above the diameter where the relativeMach number exceeds unity. This has beendemonstrated to improve efficiency by up to1% point with little change in the inducerfirst flap bending frequency.The effects of blade thickness on theperformance of the NASA 10.0 pps highpressure ratio stage are described in Ref 12,in which discrete attention was made topreserve both inducer and diffuser throatareas. Stage peak efficiencies are shown onFig 15 versus shroud inlet relative Machnumber, indicating some two percentagepoints higher efficiency with inducer tip (8.34inch diameter) blade thickness decreasedfrom .027 to .016 inch.Since typical minimum machinable inducertip blade thicknesses are close to .015 inch,blade thickness effects become significant onsmall compressors of 4.0 inch (100mm) tipdiameter or less.As mentioned previously, the radial variationof incidence at the inducer leading edge maybe small, whereas it may be desirable to tailoran incidence variation to the local Machnumber, or even select to open the throattowards the hub to provide increased throatarea. The optimization of a high Machnumber inducer design using a CFD code isbriefly discussed in Ref 13. Such codes allowindividual tailoring of the inducer andimpeller blading to suit the specificapplication needs rather than correlationwith appropriate airfoil cascadecharacteristics.

13 degBladeAngle

ENI

onicawn_■e

20 40 60 80 100% Chord

Fig 13 Impeller Blade Angle Distributions

Fig 14 Effect of Blading Types

60

40

20

0

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ency ....„.. _,....., -77-, - •;:-`)

s% 1.1 61' 1.28

0.86 1.04

0 016 inch M1ws----- 0.027

80Effici

70

60

0.60.6 0 8 1.0

Figl 7 Effect of Bleed Position

5.0 DISCUSSION.

0.1 0.15 0 20 0.25 0.30Inlet Flow Coefficient C1it12

Fig 15 Effect of Blade Thickness Ref 12.

4.3 Inducer Leading Edge Sweepback

Since research on axial turbofan stages hadshown leading edge sweepback effective as ameans of reducing shock losses , increasingblade stiffness ,and deterring buzz-saw noisegeneration, the inducer blading of theresearch compressor Fig 7, was redesigned asshown on Fig 16 ,while maintaining the samethroat area. Comparative performancecalibrations showed less than a 1% pointimpeller efficiency improvement with sweepat a relative inducer tip Mach number (onedimensional) of 1.45. The effect of shroudbleed position on this impeller is shown onFig 17 indicating that bleed at 30% chordexhibited the highest performance.

Fig 16 Swept Leading Edge Impeller.

An appraisal of design rationale for moderateto high pressure ratio centrifugalcompressors inducers has been describedbased upon lessons learned in the design,test and development of small aviation gasturbines and industrial gas compressorsduring the latter half of the 20 th century.Compressor performance improvements havemade with the introduction of transonicinducer blading featuring minimal turning upto the throat, and low thickness to pitchratios, at some expense to manufacturing costand durability.Turbomachinery aerodynamic and structuraldesign of centrifugal compressors hastransitioned from the traditional meanlineperformance characterization,streamlinecurvature internal flow analysis, and drawingboard geometry definition, to a fullyintegrated computerized procedure, fromcustomer specification to the finished out-the-door product. By the year 2000 this may beaccomplished within one month.The key to such a rapid turnover will not beentirely in the drawing creation or machinedpart, but the capability of computational fluiddynamic codes to rapidly predict the entirecompressor map, with an assurancecommitment requiring only an acceptancedelivery performance test.This indeed is the challenge, as todays mostexpensive and time consuming compressordevelopment task is performancedevelopment to satisfy design requirements.Having been exposed to the design anddevelopment of a wide diversity of specific

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speed centrifugal compressors it is apparentthat the acid test of any truly universal CFDcode, must be in the codes capability tosuccessfully predict the characteristics of allgeometries ranging from the partial emissioncompressor at the low end of the flow scale, tothe mixed flow compressor at the high flowlimit, and including fully variable geometries.Excellent agreement has been achievedbetween test and CFD code (Dawes) inducerchoke flows at inducer tip relative Machnumbers up to 1.5.The magnitude of potential improvements inefficiency, and other foreseeable reduction inlosses is generally anticipated to be minor,and it is already clear that research emphasison centrifugal compressors is being morefocused on the broadness of the efficiencyislands, stable operating flow range extension

between choke and surge, plus increasingpressure rise towards surge.

. ACKNOWLEDGEMENT.

The author wishes to acknowledge thecontributions of his past and presentcolleagues at Sundstrand Aerospace andSolar Turbines in San Diego for theirassistance, guidance,and collaboration, in themany interesting centrifugal compressorresearch and development programs that hehas been fortunate to report upon.

LIST OF REFERENCES.Ref 1. Rodgers C., "Effect of Inlet Geometryon the Performance of Small CentrifugalCompressors", AIAA-88-2812,1988.

Ref 2. Ariga L, Kasai N., Masuda S.,Watanabe Y., Watanabe L, "The Effect ofInlet Flow Distortion on the Performanceof aCentrifugal Compressor", ASME 82-GT-92,1992

Ref 3. Rodgers C., "Influence of Impeller andDiffuser Characteristics and Matchingon Radial Compressor Performance" , SAECentrifugal Compressors , 1961.

Ref 4. Rodgers C., "Centrifugal CompressorInlet Guide Vanes for Increased SurgeMargin" , ASME 90-GT-158, 1990.

Ref 5. Klassen H.K., Wood J.R., SchummanL.F., "Experimental Performance of a13.65mm Tip Diameter Tandem BladedCentrifugal Compressor Designed fora Pressure Ratio of 6.0", NASA Tech Paper1091, 1977.

Ref 6. Rodgers C.,"Small CentrifugalCompressor Tip Blockage Effects",AIAA 89-2570 ,1989.

Ref 7. Chapman D.C., " Model 250-C30/C3OB Compressor Development",AGARD -CP-282 , 1980.

Ref 8. Fisher F.B., "Application of MapWidth Enhancement Devices toTurbocharger Compressor Stages", SAE880794, 1988

Ref 9. Eynon P.A., Whitfield.A., Firth M.R.,Parkes A.J., Saxton R.," A Study of the FlowCharacteristics in the Inducer Bleed Slot of aCentrifugal Compressor". ASME 96-GT-921992.

Ref 10. Rodgers C, "The Efficiencies ofSingle Stage Centrifugal Compressorsfor Aircraft Applications". ASME 91-GT-77,1991.

Ref 11. Palmer D.L Watermann W.F.,"Design and Development of an AdvancedTwo-Stage Centrifugal Compressor", ASMEJT Vol 117, 1995.

Ref 12. G.J.Skoch., " Performance of TwolOpps Compressors with Differential Bladeand Shroud Thicknesses Operating Over a

Range of Reynolds Numbers",AIAA -87- 1745, 1987.

Ref 13. ICalogiannis.M.H.,Leblanc.A.D.,Przybytkowski.S.M., "Advanced Small HighPressure Ratio Centrifugal Compressor".AGARD -CP-537 1994.

Ref 14. Wright.P.L,Miller.D.C., "AnImproved Compressor PerformancePrediction Moder,Inst Mech EngrsC423/028,1991.

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