C. Xu_2005 - Development and Design of a Centrifugal Compressor Volute

International Journal of Rotating Machinery 2005:3, 190196c 2005 C. Xu and M. Muller

Development and Design of a CentrifugalCompressor Volute

Cheng XuIngersoll-Rand, Davidson, NC 28036, USAEmail: cheng [email protected]

Michael MullerIngersoll-Rand, Davidson, NC 28036, USAEmail: [email protected]

Received 25 April 2005

The volute is one of the key components of a centrifugal compressor. The design of the volute not only impacts the compressoreciency but also influences the operating range. The detailed flow simulation presented here helps to better understand the voluteflowmechanisms and provide design guidance in volute design to meet performance goals. In this study, the viscous Navier-Stokesequations are used to simulate the flow inside the vaneless diuser and volute. The detailed flow structures for dierent volutetongue geometries are studied in detail. The numerical calculations are compared with experimental data and good agreement isfound. The results also suggest direction for further investigations in volute design.

Keywords and phrases: centrifugal compressor, volute, tongue structure.

1. INTRODUCTION

With the development of aviation and industry, compressiontechnology has made large progress in both axial and radialcompressors. Centrifugal compressors can make use of theradius change across the impeller. This allows for more dif-fusion within the impeller blade; thus a centrifugal stage canproduce much higher specific work transfer than an axialstage. At high flow, the eciency of a centrifugal compres-sor is not as high as that of an axial compressor. However,centrifugal compressors have wide application for aviationand industry, and are preferred for handling small volumet-ric flows. The physical length is shorter and the operatingrange is wider as compared to an axial compressor.

Increased eciency has been the traditional emphasis ofcompressor design. However, increased operating range hasbecome important in an age of ever-increasing productivityand energy costs; a multitasking compressor has economicbenefits.

Centrifugal compressors consist of three main parts, asshown in Figure 1. The first is a rotating impeller, which im-parts work to the gas by increasing its angular momentum.

This is an open access article distributed under the Creative CommonsAttribution License, which permits unrestricted use, distribution, andreproduction in any medium, provided the original work is properly cited.

The fluid static pressure and absolute velocity (stationaryframe of reference) increase through the impeller passage.The second component is the diuser section, often withvanes to increase the eectiveness. The diuser converts thekinetic energy into the static pressure by decelerating thefluid. The third and final component is a volute or collec-tor, used for collecting the gas from diuser and delivering tothe outlet pipe.

A volute has two functions: collection and diusion. Thevolute must collect and transport the fluid to the down-stream system. It also raises the static pressure by convertingkinetic energy (u2) to potential energy (static pressure). Thelatter function has performance benefits, as the dischargepressure is increased.

The fluid leaving the impeller follows a logarithmic spi-ral due to the rotation of the impeller. This strong swirl of thefluid entering the volute complicates the volute design, as theflow is strongly three-dimensional. Diusion in the volute ismore eective for the kinetic energy associated with the tan-gential velocity; this is accomplished by increasing the cross-sectional area from volute tongue to exit, creating a conicaldiuser. But there is often an eciency loss arising from theinability of the volute to convert the kinetic energy associatedwith the radial component of the velocity out of the diuser.This eect is compounded when a vaned diuser is used, asthe vanes create amore radial flow into the diuser. Typically,

Development and Design of a Centrifugal Compressor Volute 191

(a) (b)

Figure 1: Centrifugal compressor layout. (a) A cross-sectional view,showing the impeller followed by a vaned diuser, and a volute. (b)An isometric view of the package.

the tradeo between increased recovery of the vaned diuserand reduced recovery of the volute suggests using a vaneddiuser due to the performance requirements.

Little research has been done to understand the flow involutes, especially in the tongue area (Ayder [1], Ayder andVan den Braembussche [2], Ayder et al. [3], and Gu et al. [4]).Recently, research (Ayder [1]) has found that a good volutedesign could improve the compressor performance and oper-ating range. Most volute studies have been performed on ex-isting geometries and can be categorized into two groups: vo-lute flow analysis (Ayder [1]) and flow interactions betweenvolute and impeller (Gu et al. [4]). Most of the studies werebased on the pump. There is very limited literature to studythe centrifugal compressor volute. There is little informationavailable for the guidance of the volute design and design pa-rameters influences of the performance.

There are five key geometrical parameters in volutedesign (Ayder [1]) circumferential variation of the cross-sectional area; shape of the cross-section; radial position ofthe cross section; position of the volute inlet; and tongue ge-ometry. A design requirement is to make the circumferentialpressure distribution uniform.

The tongue of a volute is located at 360 point of thevolute, corresponding to a joint of the volute smallest area,largest area, and exit cone. At this point, the flow in the voluteencounters a dividing plane, the so-called tongue, abovewhich the flow exits the compressor, below which the flow re-enters the volute. The leading edge of the tongue is analogousto that of an airfoil, a sharp leading edge has low blockage butis sensitive to angle of attack; while a well-rounded leadingedge has large blockage but is insensitive to incidence. Thisdierence manifests itself as a design tradeo: a sharp volutetongue provides the absolute peak eciency, while a volutewith a round tongue has better o-design performance com-pared.

Given the economic incentive for both eciency andrange, the focus of this paper is to gain a better understand-ing of the influence of the tongue. Given this knowledge, acompressor with higher eciency and greater stable operat-ing region can be designed.

Traditional volute design is based on two-dimensionalanalysis, and the emphasis is on collection and less on thediusion function. However, with the advent of advancedfluid modeling tools, it is possible to design a volute usingthree-dimensional analysis. The volute design and analysispresented is indeed part of a new product in developmentat Ingersoll-Rand.

2. VOLUTE DESIGN CONSIDERATIONSAND PROCESS

The volute design objective is to meet the overall compres-sor eciency and operating range goals. The cross-sectionalshape is designed mainly from manufacturing and com-pressor packaging constraints. Volutes are typically cast, andare therefore expensive to manufacture and not practical tochange. First-pass designs must achieve a high success, and todo this the fundamental flow structure must be understoodalong with manufacturing limitations. In this study, the ve-locity enters the volute at a tangential angle of about 65 atdesign flow, thus the dominant part (> 80%) of the total ki-netic energy available is associated with velocity in the radialdirection.

The volute design process in this study embraces twobasic considerations. First, angular momentum and massconservation are assumed. Second, during the first designiteration, it is assumed that the pressure and velocity areindependent of the angular angle, and the tongue leakageis assumed. Based on these considerations, the volute sec-tion at a given angle can be defined. Dierent volute cross-section shapes are analyzed. It is found that there are onlysmall dierences that occur for all the asymmetric volute sec-tions. For this particular case, the installation space is unre-stricted, so the volute radial position is chosen by applyingangular momentum conservation. The position of the vo-lute inlet is designed tangential to the diuser end wall asshown in Figure 1. Ease of manufacture drives this consider-ation.

An important design parameter is the tongue geometry.Most of the recent research (Chu et al. [5] and Dong [6])investigates the eect of the tongue shape on head, pressurefluctuations, and noise in a centrifugal pump. Irabu et al. [7]study the water flow around the volute tongue using a par-tial tracking method. The research is focused on the low flowoperating region. The study finds that the flow separationoccurs from both upper and lower surfaces of the tongue.It is important to understand the flow structure around thetongue for improving the design.

After the basic volute shape is constructed, a three-dimensional turbulence commercial software CFX-5 solveris used as to analyze the volute performance. The volutepressure recovery and internal losses are calculated with the

192 International Journal of Rotating Machinery

Figure 2: Ingersoll-Rand test rig of the single-stage compressor.The inlet pipe is the white pipe coming from the left of the image.The compressor discharges through the flex hose visible just abovethe inlet pipe. The regulating butterfly valve can be seen in the back-ground.

three-dimensional analysis at design flow rate. Some finalmodifications are made after this initial analysis, replacingthe previous assumptions with analytic values. Dierenttongue geometries are analyzed. From these analyses, an op-timal radius is found to provide a good eciency in a wideoperating range.

3. EXPERIMENTAL STUDY

The final volute design was installed with its correspond-ing compressor and tested at the Ingersoll-Rand Develop-ment Laboratory. The test and data acquisition proceduresare based on the ASME performance test procedure, ASME[8].

The test compressor consists of an inlet pipe, impeller,vaneless diuser, vaned diuser, vaneless diuser, and vo-lute. This study considers the volute preceded by a low so-lidity vaned diuser as shown in Figure 1. The test circuit isan open loop testing system as shown in Figure 2. The flowrate is regulated by a butterfly valve located at the dischargepipe. The performance characteristic was tested through fivetotal temperature and total pressure measurements at inletand discharge of the volute exit cone. For determining thestatic pressure ratio, five pressure transducer taps, circum-ferentially distributed, were mounted on the wall at the dis-charge of volute exit cone. The mass flow was measured byusing an ASME [8] nozzle located at the end of dischargepipe. The tip clearance of the impeller at tip was measuredwith graphite pins. The test uncertainty for the flow rateis less than 2% at 95% confidence level. The head and ef-ficiency uncertainties were kept under 2.2% and 2.5%, re-spectively, with the same confidence level based on the sys-tem uncertainty analysis (Kline [9]). The performance testresults are shown in Figure 3. It is shown that the compres-sor developed by this research has a wide operating rangeup to 40% with corresponding high eciency over a widerange.

0.9

0.85

0.8

0.75

0.7

0.65

0.6

0.55

0.5

0.45

0.4

0.35

,

0.6 0.8 1 1.2 1.4

/0

CalculatedCalculated

ExperimentExperiment

Figure 3: Calculated and measured performance characteristics,showing a good agreement between the two.

Figure 4: The computational mesh, which includes the impeller,vaned and diuser sections, and the volute.

4. MATHEMATICALMODEL DESCRIPTIONS

The volute, together with the impeller and vaned diuser,is computationally investigated using a commercial Navier-Stokes solver, CFX 5.6. This code is a fully 3D viscous turbu-lent flow solver. The geometrical discretization of the com-pressor stage is made for the computational analysis.

Structure hexahedral cells are generated in the impellerinlet, impeller, and diuser zones. An unstructured tetrahe-dral mesh with prism layers near the wall surfaces is used forthe volute. The mesh is shown in Figure 4. The mesh sizes forinlet pipe, impeller, diuser, and volute are 2.2 104, 5.2 105, 2.4105, and 5.7105 nodes, respectively. In the volutemesh generation, a refine zone is defined near the tongue.The frozen-rotor method is used at interfaces between rotat-ing and stationary elements. Frozen rotor indicates that therelative orientation of the two interface components acrossthe interface is fixed. The calculation of the downstream sur-face of the interface plan was based on the average mixingplane approach. Interfaces are located at the surface betweenthe inlet and impeller, and between the impeller and diuser.


Tongue center

(a)

Tongue center

(b)

Figure 5: The tongue structures, showing a variation in the leading edge. (a) Sharp tongue, r/b = 0 (knife-edge). (b) Round tongue,r/b = 0.3.

The turbulence model used here is based on a zero-equation model; only an algebraic equation was used to cal-culate the viscous contribution from turbulent eddies.

The boundary conditions at the inlet are enforced by us-ing total pressure and total temperature. The flow to the in-let pipe is assumed normal to the inlet surface. The outletboundary condition is applied by using a variable static pres-sure proportional to the kinetic energy at the outlet. The flowrate of the compressor is changed by modifying the staticpressure to kinetic energy ratio at the outlet condition, whichsimulates increasing backpressure. Zero-slip boundary con-ditions were imposed over the impeller blades, diuser air-foils, and all solid walls.

The calculation was performed using an HP 8000c work-station. The time step is set to 5.0 105 seconds for all cal-culations. The calculation is assumed to be convergent whenthe ratio between the sum of the residuals and the sum ofthe fluxes for a given variable in all of the cells is reduced tofive orders ofmagnitude. Intensive grid-size dependence testswere carried out [10] and overall compressor performancewas compared with dierent mesh sizes. The mesh is deter-mined to be fine enough when with the flow rate variationsless than 5% with the increased mesh density.

5. RESULTS ANDDISCUSSION

In this study, four dierent types of tongues are investigated,that is, r/b2 = 0, r/b2 = 0.75, r/b2 = 0.15, and r/b2 = 0.3.Figure 5 shows two typical tongue structures considered,sharp tongue (r/b2 = 0) and round tongue (r/b2 = 0.3),showing the variation at the leading edge. It is worth topoint out that all tongue locations studies here are of cut-back tongues, which allow some flow recirculation. The cutback is about 9 from the full tongue location. Three oper-ating points are calculated for each tongue, while five op-erating points are calculated for the final designed volute.The calculated performance based on r/b2 = 0.3 tongue is

shown in Figure 3. It shows that the calculation results arein good agreement with experimental data. It is shown thatonly at the operating point near choke, calculation showedmore choke margin than experimental data.

Figures 6 and 7 show the secondary flow vectors (i.e., nor-mal to the mean flow) at sections = 180 and = 360, forboth the sharp and round tongues. It can be seen that theflow in the volute sections has a single vortex structure, asopposed to two counter-rotating vortices. The round tonguecreates significant blockage near the tongue. This blockageforces secondary flow center away from the tongue area. Fig-ure 8 shows the similarity in the basic flow structure near thetongue.

Figures 9 and 10 show computational static pressure con-tours for the two tongue geometries. It is shown that sharptongue has a more uniform pressure. This may be due tothe impacts of the tongue blockage. Figure 11 shows the to-tal pressure Contours for both sharp and round tongues.This figure indicates that the round tongue produces a highloss near the tongue area. The loss associated with the ob-served asymmetry may produce the asymmetric secondaryflow profile seen in Figures 6 and 7.

Figure 12 shows the pressure coecient as a function ofthe flow coecient for the two dierent volute tongue ge-ometries. This figure shows that the round tongue providesthe improved pressure coecient at high flow as comparedto the sharp tongue. However, the absolute peak pressure re-covery is attained with the sharp tongue, at a loss of range.

6. CONCLUSION

In this study, a detailed CFD analysis for dierent volutetongues is presented. The detailed flow structures are shownand the compressor performance is calculated for dierenttongues. A performance test validates the CFD analysis forthe final design. It is shown that the round tongue pro-duces better performance than the sharp tongue. It is also


(a) (b)

Figure 6: The velocity vectors versus the volute exit ( = 180). A single vortex is observed, indicating a well-sized volute. (a) Sharp tongue.(b) Round tongue, showing a shifted vortex center resulting from increased blockage near the tongue.

(a) (b)

Figure 7: The velocity vectors at the volute exit ( = 360). Similar observations as in Figure 6 are made. (a) Sharp tongue. (b) Roundtongue.

Tongue

(a)

Tongue

(b)

Figure 8: The velocity vectors near the intersection between the midplanes of the diuser and tongue, showing the basic flow structure.(a) Sharp tongue. (b) Round tongue.


Pressure

[Pa]

350 000349 786349 571349 357349 143348 929348 714348 500348 286348 071347 857347 643347 429347 214347 000

(a)

Pressure

[Pa]

350 000349 786349 571349 357349 143348 929348 714348 500348 286348 071347 857347 643347 429347 214347 000

(b)

Figure 9: Static pressure contours versus the volute exit ( = 180). (a) Sharp tongue. (b) Round tongue, indicating a much less uniformpressure distribution.

Pressure

[Pa]

349 000348 786348 571348 357348 143347 929347 714347 500347 286347 071346 857346 643346 429346 214346 000

(a)

Pressure

[Pa]

349 000348 786348 571348 357348 143347 929347 714347 500347 286347 071346 857346 643346 429346 214346 000

(b)

Figure 10: Static pressure contours at volute exit ( = 360), showing similar results as in Figure 9. (a) Sharp tongue. (b) Round tongue.

Total pressure

[Pa]

350 300350 250

350 200

350 150350 100

350 050

350 000349 950

349 900

349 850

349 800349 750349 700

349 650349 600

(a)

Total pressure

[Pa]

350 300350 250350 200350 150350 100350 050350 000349 950349 900349 850349 800349 750349 700349 650349 600

(b)

Figure 11: Total pressure contour at the discharge of exit cone. (a) Sharp tongue. (b) Round tongue, showing a much steeper pressuregradient, contributing to the secondary flow seen in Figure 7.


0.6

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0.6 0.8 1 1.2 1.4

/0

r/b2 = 0.3r/b2 = 0

Figure 12: The performance characteristics of the two volute con-figurations discussed. The tradeo between peak performance ofthe sharp tongue and the greater range of the round tongue is evi-dent.

important that the round tongue is easier to cast, an impor-tant manufacturing consideration. This study sets a generalguideline for centrifugal volute design.

ACKNOWLEDGMENTS

The authors would like to thank the Ingersoll-Rand Corpora-tion, Air SolutionsHeavy Industrial Divisionfor alloca-tion of resources, and the engineers in the Development Labfor their time and support in instrumentation and collectionof experimental data.

REFERENCES

[1] E. Ayder, Experimental and numerical analysis of the flowin centrifugal compressor and pump volutes, Ph.D. disserta-tion, Von Karman Institute for Fluid Dynamics, Rhode SaintGenese, Belgium, April 1993.

[2] E. Ayder and R. Van den Braembussche, Numerical analy-sis of the three-dimensional swirling flow in centrifugal com-pressor volutes, Transactions of the ASME, Journal of Turbo-machinery, vol. 116, no. 3, pp. 462468, 1994.

[3] E. Ayder, R. Van den Braembussche, and J. J. Brasz, Exper-imental and theoretical analysis of the flow in a centrifugalcompressor volute, Transactions of the ASME, Journal of Tur-bomachinery, vol. 115, no. 3, pp. 582589, 1993.

[4] F. Gu, A. Engeda, M. Cave, and J.-L. Di Liberti, A numericalinvestigation on the volute/diuser interaction due to the ax-ial distortion at the impeller exit, Transactions of the ASME,Journal of Fluids Engineering, vol. 123, no. 3, pp. 475483,2001.

[5] S. Chu, R. Dong, and J. Katz, Relationship between unsteadyflow, pressure fluctuations and noise in a centrifugal pump-part B: eects of blade-tongue interactions, Transactions ofthe ASME, Journal of Fluids Engineering, vol. 117, pp. 3035,September 1995.

[6] R. Dong, S. Chu, and J. Katz, Eect of modification to tongueand impeller geometry on unsteady flow, pressure fluctua-

tions and noise in a centrifugal pump, Transactions of theASME, Journal of Turbomachinery, vol. 119, no. 3, pp. 506515, 1997.

[7] K. Irabu, E. Tamazato, and I. Teruya, Velocity measurementof flow around a volute tongue of the vaneless radial diuser,in Proc. 3rd Asian Symposium on Visualization (ASV 94),Chiba, Japan, May 1994.

[8] ASME, Performance Test Code on Compressors and Exhausters,PTC10-1997, 1997.

[9] S. J. Kline, The Purpose of uncertainty analysis, Transactionsof the ASME, Journal of Fluids Engineering, vol. 107, pp. 153160, June 1985.

[10] C. Xu and R. S. Amano, A hybrid numerical procedure forcascade flow analysis, Numerical Heat Transfer: Part B: Fun-damentals, vol. 37, no. 2, pp. 141164, 2000.

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