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Forced Response Predictions in Modern Centrifugal Compressor Design - A Progress Report Presented at 10th National Turbine Engine HCF Conference -

C. Smythe

MIT Gas Turbine Laboratory

J. Liu Honeywell Engines, Systems, & Services

C. S. Tan

MIT Gas Turbine Laboratory

1.0 Motivation and Background

Unsteady aerodynamic phenomena, acting as a source of excitation of blades, are a key element in the prediction of high cycle fatigue (HCF). It is well-known that air loads in turbomachines, axial as well as centrifugal, are inherently unsteady; however, discussions at an AFOSR workshop in 1995 [1] emphasize that it remains unclear if all excitation mechanisms have been identified. The complexity of the flow features in a modern, high performance centrifugal compressor stage (for aircraft turbine engines) raises the following question: What key physical flow effects must be included in an analytical/computational model in order to obtain a useful and adequate characterization and prediction of the unsteady loads that can lead to aeromechanic difficulties? Thus an objective of this work is to evaluate a system of CFD (MSU Turbo [2]) and structural dynamics analysis (ANSYS) as a tool for forced response predictions in modern centrifugal compressor design. Often aeromechanic difficulty arises out of design steps taken to raise the performance level. Therefore, the issue of sensitivity of the physical flow effects that impact aeromechanic behavior to incremental design changes (made to achieve centrifugal compressor stage performance enhancement) is of engineering significance. An evaluation of the predictive capability of MSU Turbo in combination with ANSYS for forced response and an interrogation of the flow effects is performed using two nearly identical centrifugal compressor stage configurations. The enhanced design with improved performance is similar to the production design, the most notable aerodynamic difference being that the enhanced design has a smaller impeller-diffuser gap than the production design; however, the two designs exhibit distinctly different aeromechanic behaviors. 2.0 Scope of Paper and Statement of Key Results

This paper describes the progress towards evaluating the system of tools consisting of MSU Turbo and ANSYS for predicting the forced responses observed in the production and the enhanced designs alluded to in Section 1.0, focusing on the implications of the unsteady aerodynamics results from CFD analysis. The computations that have been implemented on both designs serve to generate (unsteady) flow fields from which the following are extracted: (i) the corrected speedline at 96.2% for the enhanced design; (ii) the corrected speedlines at 96.2% and 102% for the production design; (iii) the unsteady loading characteristic on the impeller main and splitter blades for analysis and to be used as the forcing function for the follow-up ANSYS calculations;

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(iv) flow events that drive the observed performance changes in the enhanced and the production design. The computed results from (i) to (iii) have been assessed against the data from the aerodynamic test rig and will be assessed against the strain gauge data. The computed results from (i) and (ii) were also compared with the flow fields previously computed using ADPAC [3] based on a blade count of 16 impeller main blades, 16 splitter blades and 24 diffuser vanes instead of the actual blade count of 17 impeller main blades, 17 splitter blades and 25 diffuser vanes. The modified blade count was chosen for implementation in ADPAC to avoid modeling the entire impeller wheel and the entire diffuser so as to diminish the requirements for computational resources.

Overall, the results from MSU Turbo appear to be in accord with both

aerodynamic and structural test data on a preliminary basis (this comparison is beyond the scope of this paper and will be reported in further detail in future works). In addition, the following specific observations can be inferred from the computed results: (1) unsteady pressure waves emanating from the diffuser leading edge plane appear to be stronger than in the enhanced design, and the pressure wave pattern of the two designs differs; (ii) the diffuser pressure recovery for both the enhanced and production designs are similar; (iii) the computed unsteady loading is confined to the last 10 to 15% of the impeller main and splitter blades and the level of unsteadiness increases toward the hub region; (iv) the level of computed unsteadiness appears stronger on the impeller pressure surface and on the splitter suction surface; (v) the fundamental vane passing frequency is dominant ; (vi) there appears to be a distinct difference between the response of the splitter blades in the enhanced design and that of the splitter blades in the production design.

This paper is organized as follows: Section 3.0 presents the key results from post-

processing of unsteady computed flow fields in the two nearly identical centrifugal compressor stage designs to support the observations stated above. An overall summary is given in Section 4.0. 3.0 Results

Representative results are presented to elucidate the effect of marginal change in impeller-diffuser gap (from the production to the enhanced design) on time-averaged compressor performance and on the unsteady loading characteristics of impeller main and splitter blades. The characterization of unsteady loading on impeller main and splitter blades should be viewed as preliminary.

3.1 Characterization of Computed Time-average Performance MSU Turbo has been implemented to obtain several sets of unsteady three-dimensional flow in the impeller-diffuser stage of both the production and the enhanced design. These sets of computed results have been post-processed to obtain the performance characteristics at a corrected speed of 96.2 % for both the production and the enhanced design as well as at a corrected speed of 102 % for the production design. These computed performance characteristics are shown in Figure 1. Both the corrected mass flow (abscissa) and the pressure ratio (ordinate) have been appropriately normalized. The results in Figure 1 show that for the range of computed corrected mass

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flow the pressure ratio of the enhanced design is higher than that of the production design; this observation is in accord with the design intent as well as with the test data. The time-averaged performance at a similar normalized corrected mass flow for the two designs at 96.2 % corrected speed is tabulated in Table 1. It is seen that the diffuser pressure recovery in the two designs are nearly identical, though the computed loss in stagnation pressure is higher in the diffuser passage of the production design (i.e. flow in the production has more loss). The impeller of the enhanced design has a higher (adiabatic) efficiency than that of the production design by about 2 %. 3.2 Unsteady Aspects of Computed Flow The sets of computed flow alluded to above have not been exhaustively post-processed and interrogated to identify the flow process/processes responsible for the observed change in performance (see above) and the observed difference in forced response (see below) in the two designs. Nevertheless certain aspects of the unsteadiness have been examined and these are described in this section. Selected results based on CFD results from ADPAC [3] are presented first and these are shown in Figures 2 to 4. Figure 2 constitutes a snapshot of unsteady Mach number contours (Figure 2a) and unsteady static pressure contours (Figure 2b) at one time instant. These contours were taken at the mid-span plane from upstream of the impeller to downstream of the diffuser for both the production and enhanced design. It is deduced from these sets of time-varying Mach number and static pressure contours that unsteady pressure waves emanating from the diffuser leading edge plane are stronger and of a different pattern in the enhanced design than in the production design. The blade static pressure distributions corresponding to maxima (red), time-averaged (green) and minima (blue) are shown in Figure 3a for the impeller main blade and Figure 3b for the splitter blade. It can be inferred from Figure 3 that the unsteady loading is confined to the last 10 to 15% of the impeller main and splitter blades and that the level of unsteadiness increases toward the hub region; it is also seen that the level of unsteadiness appears stronger on the impeller main and splitter blade pressure surfaces, though there is appreciable unsteadiness on the splitter suction surface. Figure 4 shows the time variation of static pressure at 95% meridional distance for three spanwise locations (10%, 50% and 90% span) on the splitter blade of the enhanced design and the corresponding power spectrum analysis. The results in Figure 4 again reflect the inference made above. Power spectrum analysis, shown only for the pressure surface of the enhanced splitter blade pressure surface, of unsteady loading on the impeller main and splitter blades indicates that the 3rd harmonic of the 3 diffuser vane computational sector corresponds to fundamental vane passing excitation frequency in the ADPAC solutions. Selected results from MSU Turbo solutions (at a normalized corrected mass flow of 1.018) are shown in Figure 5 to 7. Figure 5 and 6 show respectively the impeller exit static pressure contours and entropy contours at three time instants over a period of time variation in these variables. One might deduce that the difference in the static pressure contour distribution in the two designs implies a difference in the unsteady loads in the trailing edge region of the splitter and impeller main blades (this aspect will further be assessed in the future for its implication in the observed difference in forced response

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between the two designs). One would infer from the impeller exit entropy distribution contours in Figure 6 that the production design appears to have a higher level of entropy production. This decrease in entropy production from the production to the enhanced design is reflected in the improvement in the enhanced impeller performance over that of the production impeller. In an attempt to rationalize the difference in the computed compression wave pattern from the diffuser into the impeller trailing edge region, we examine the (time-averaged) flow incidence angle distribution into the diffuser vane passage. This is shown in Figure 7 at three spanwise locations (10%, 50% and 90% of span) on the leading edge plane of the diffuser. While there is a difference in the time-averaged incidence angle distribution for the two designs, a connection between this difference and the observed difference in compression wave pattern propagating upstream into the impeller blade region cannot yet be made. If this connection can be established, then a design guideline to avoid forced response difficulty in the enhanced design can be put forward. 4.0 Overall Summary

MSU Turbo has been used successfully (contingent upon the availability of good initial flow conditions to ensure robust implementation of MSU Turbo) to generate performance characteristics of two nearly identical centrifugal compressor designs referred to as production and enhanced design. The observed improvement in the computed time-averaged performance of the impeller for the enhanced design with a smaller impeller-diffuser gap is in accord with the results of Shum [4] and Murray [5]; this is also in accord with the aerodynamic design intent. It is suggested that the observed difference in the forced response for the two designs could be associated with the difference in the impeller exit flow delivered to the diffuser; this difference in impeller exit flow results in a difference in the diffuser inlet flow incidence angle and hence in the compression wave pattern into the impeller blade region.

Based on preliminary computed results (not reported here), a system of tools consisting of the CFD code MSU Turbo and the structural analysis code ANSYS appears to be a promising predictive scheme for forced response in modern centrifugal compressor design.

5.0 References 1. AFOSR Workshop. Basic research issues in aerodynamics, structure dynamics and

control of high cycle fatigue. MIT Gas Turbine Laboratory, October 1995. 2. Chen, J. P., Briley, W.R., “A Parallel Flow Solver for Unsteady Multiple Blade Row

Turbomachinery Simulations,” Proceedings of ASME TURBO Expo 2001, New Orleans, LA, June 4-7, 2001.

3. Mansour, M., Kruse, M., “Time Unsteady Impeller-Diffuser Interaction in Gas Turbine Engines,” Proceedings of the 5th National Turbine Engine High Cycle Fatigue (HCF) Conference, Chandler, AZ, March 7, 2000.

4. Shum, Y.K., 2000, “Impeller-Diffuser Interaction in Centrifugal Compressors,” PhD Thesis, Massachusetts Institute of Technology.

5. Murray, N.P., 2003, “Effects of Impeller-Diffuser Interaction on Centrifugal Compressor Performance,” Master Thesis, Massachusetts Institute of Technology.

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Acknowledgement This research has been sponsored by GUIDE III Consortium with Dr. C. Cross as Contract Monitor and with Dr. J. Griffins as subcontract monitor via Carnegie -Mellon University. We are grateful to the following people that provide help in enabling the successful execution of MSU Turbo on the MIT Gas Turbine Laboratory Computer System: Dr. J. P. Chen and his colleagues of MSU, Dr. S. Gorrell and Mr. D. Car of CARL AFRL, Mr. R. Haimes of MIT ACDL and Mr. P. Warren of MIT Gas Turbine Laboratory. We are also grateful to Dr. M. Mansour of Honeywell ES&S for his contributions to the ADPAC analysis and to Dr. J. Panovsky for his advice and for facilitating the entire project. 6.0 Figures and Tables

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0.92 0.94 0.96 0.98 1 1.02 1.04 1.06 1.08 1.1

normalized mcorr

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Enhanced96%

Production96%

Production102%

Figure 1 - Computed Speedlines using TURBO

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Figure 2a - Absolute mach contours at 50% span for the Production at 102% speed and Enhanced at 96.2% speed – Results from ADPAC

Figure 2b - Static Pressure Contours at 50% for the production at 102% and Enhanced at 96.2% Speed - Results from ADPAC

Production Enhanced

Mach contours range from 0 to 1.34

Production Enhanced

Pressure contours range from 0 to 3.3

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Figure 3a - Unsteady static pressures on the impeller main blade for the Production at 102% and Enhanced at 96.2% speed

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Pressure Surface Suction SurfaceT.E. L.E. T.E.

% Meridional Distance

Figure 3b - Unsteady Static pressures on the impeller splitter blade for the Production at 102% and Enhanced at 96.2% speed

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0 1000 2000 3000 4000 Figure 4 - Static pressure vs. time and power spectrum analysis for the enhanced splitter blade pressure surface at 95% meridional distance

Production

Time = to

Time = 0.8T

Time = 0.4T

Enhanced

Time = to

Time = 0.4T

Time = 0.8T Figure 5 - TURBO results: Impeller exit plane static pressure contours at 3 time instants within one period, T, for the Enhanced and Production designs at 96.2% speed

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Production

Time = to

Time = 0.8T

Time = 0.4T

Enhanced

Time = to

Time = 0.4T

Time = 0.8T Figure 6 – TURBO results: Entropy contours at impeller exit plane at 3 time instants within one period, T, for the Production and Enhnaced designs at 96.2% speed.

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pitchwise distance

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Prod 10% SpanEnh 10% SpanProd 50% SpanEnh 50% SpanProd 90% SpanEnh 90% Span

Figure 7 - Diffuser inlet incidence angle for the Production and Enhanced Designs at 96.2% speed

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Production 96.2% Enhanced 96.2%

Normalized Cp diffuser 1.0 1.0

p t impeller (normalized) 1.04 1.06

p t vaneless space (normalized) 0.21 0.21

p t diffuser (normalized) 0.22 0.21

p t stage (normalized) 0.95 0.98

Impeller efficiency (normalized) 1.0 1.02

Table 1 - Normalized Pressure Recovery, Pressure Rise, and Efficiency for the Production and Enhanced designs at 96.2% speed