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02

On the choice of impeller exit blade angles in the spanwise direction of a high-pressure

ratio centrifugal compressor

Dr. Justin Jongsik Oh

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02Motivation & Objectives

In general, the impeller exit blade angle transferred from the meanline design to a three-dimensional design step is set constant from the hub to the shroud, but can be varied as long as therms (root-mean-square) angle stays unchanged from the meanline value.

The same transonic centrifugal compressor as previous studies in series [1][2][3] was taken in thepresent study where the shroud was 4.72° unloaded relative to the hub for the rms angle of 52.3°(*)at the impeller exit.

It looks there was a special reason why they did not keep the angle constant from the hub to theshroud with 52.3°. Total 5 cases of blade angle profiles were designed while keeping the rms angleunchanged, including the original one, to investigate how both impeller and compressoraerodynamic performance change.

Through the same CFD numerical analysis, their effects on aerodynamic performance wereinvestigated.

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(*) measured from a tangential reference, which is a preferred convention in centrifugal compressors and pumps

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02Transonic Centrifugal Compressor

Modified design of a 6:1 pressure ratio centrifugal compressor

Impeller d2 = 161 mm

Impeller b2 = 5.16 mm

Specific speed (Ns) = 98

Corrected flow = 1.033 kg/s

Corrected speed = 68384 rpm

Machrel,1t =1.18

Channel-wedge diffuser

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[4]

Baseline

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02Numerical Methods

CNSTURBO, turbomachinery CFD code developed by author since 1996Time marching 3D steady-state FVM

K-omega turbulence model

Artificial dissipation of 2nd/4th-order

Multi-grid acceleration

Residual smoothing

Rotor tip clearances

Mixing plane interface

Structured multi-block grids

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02Impeller Blade (Camberline) Angle Distributions

Total 5 cases of different angle distributions from the hub to the shroud at the impeller exit

A linear distribution of the angle at the impeller exit between the hub and the shroud

Minimum changes of meridional angle distributions were tried, but they look required to retain a smooth transition.

As a result, both lean and wrap angles were changed (as seen from Theta(θ) in the table).

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2E52E4Baseline,2E3,2E2, Baseline

2E32E2

2E4

2E5

Hub

Shroud

(*) measured from a tangential reference, which is a preferred convention in centrifugal compressors and pumps

(*)

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02Impeller blade shapes

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Case 2E2 is an exact reverse of the original design.

Case 2E3 is a constant angle from the hub to the shroud.

Case 2E4 is when the shroud is 8° unloaded.

Case 2E5 is when the shroud is 12° unloaded.

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02Impeller & Compressor Overall Performance (at 100% design speed)

The worst performance in both impeller and compressor was found at Case 2E2 (which is an exact reverse).

The highest PR and accordingly the highest flow in both impeller and compressor were found at Case 2E5.

The highest design efficiency in both impeller and compressor was obtained at Baseline.

No big differences in the surge margin were found among Baseline, Case 2E4 and Case 2E5.

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02Deep dive into Baseline, 2E2 and 2E5

Why does Baseline show the highest efficiency near design flow ?

Why does Case 2E5 show the highest PR ?

Why does Case 2E2 show the lowest performance ?

In order to find answers, total 6 cross sections were taken to see the development of secondary flows in the impeller passage. The same approach of my previous study [5] was applied using the normalized relative helicity contours which would be the best in understanding the structure of secondary vortices.

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From Reference[5],

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02Overall developments of secondary flows

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No big differences in the vorticity structure up to Section (III)

Clear differences in the main passage vortex observed at Section (VI).

Splitter

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02Secondary flows near impeller exit

Case 2E2 starts to lose the strength of the main passage vortex, augmented by shroud loadings, from Section (V), leading to the lowest PR and efficiency over the range.

Case 2E5 shows a similar structure to Baseline up to Section (V), but loses the passage vortex at Section (VI) like Case 2E2.

Baseline keeps the passage vortex more stretching over the pitch, especially in the passage between splitter suction and full-blade pressure sides.

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SplitterSplitter

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02Summary

Among total 5 cases of different spanwise distributions of the impeller exit blade angle, it was confirmedthat Baseline (i.e., the original design) shows the best performance in terms of compressor pressure ratioand efficiency near design point.

Case 2E2, with an exact reverse angle distribution relative to Baseline, shows the worst performance. Itlooks that performance degradation is caused by a weaker passage vortex developing at the impeller exit,triggered by the unloaded hub.

Case 2E5 shows the highest pressure ratio, thanks to the highly-loaded hub, but due to too much unloadedshroud, compressor efficiency failed to keep the highest near design point.

When both hub and shroud are simply set constant as a meanline exit angle (β2b), like Case 2E3,compressor performance may not be an optimum.

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02References

[1]

[2]

[3]

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/Paper/Online/ of TurboAeroDesign.com

/Paper/Online/ of TurboAeroDesign.com

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02References

[4]

[5]

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