numerical study on the range enhancement of a centrifugal ... · numerical study on the range...

Journal of Mechanical Science and Technology 26 (5) (2012) 1371~1378

www.springerlink.com/content/1738-494x DOI 10.1007/s12206-012-0320-z

Numerical study on the range enhancement of a centrifugal compressor

with a ring groove system† Chi-Yong Park1, Young-Seok Choi2,*, Kyoung-Yong Lee2 and Joon-Yong Yoon3

1Advanced Engineering Team, Halla Climate Control Corp., 1689-1 Sinil-dong, Daedeok-gu, Daejeon, Korea 2Green Energy System Technology Center, Korea Institute of Industrial Technology, Cheonan, Korea

3Department of Mechanical Engineering, Hanyang University, Ansan, Korea

(Manuscript Received October 10, 2011; Revised January 20, 2012; Accepted January 30, 2012)

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Abstract A numerical study of casing treatments on a centrifugal compressor to improve stability and stall margin is presented. High efficiency,

high pressure ratio, and a wide operating range are required for a high-performance centrifugal compressor. A ring groove casing treat-ment is effective for flow range enhancement in centrifugal compressors. Compressor performance was analyzed according to the ring groove location, and the results were compared with the case without a ring groove. The effect of guide vanes in the ring groove was also investigated. Four variants of grooves were modeled and simulated using computational fluid dynamics to optimize the groove location. Numerical analysis was performed using a commercial code ANSYS-CFX program. The simulation results showed that the ring groove increased the operating range of the compressor. The ring groove with guide vanes improved both performance of the compressor at low flow rates and the stall margin of the compressor.

Keywords: Centrifugal compressor; Ring groove; Stall margin; Computational fluid dynamics ---------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------------- 1. Introduction

Centrifugal compressors are extensively used and have the advantages of a compact design and a high pressure ratio per stage. The pressure ratio and efficiency at the design point are the most important factors among the various performance indices. The operating range, depending on choking and surge, is also very important. For example, the operating range of compressors in marine turbochargers is especially significant to manufacturers. With a wide operating range, the manufac-turing cost could be reduced by decreasing the number of series models to obtain the required flow range. Improving the operating range makes the compressor more stable and more competitive. Various methods have been studied to improve the compressor surge margin experimentally and numerically. Ring groove casing treatments have been effectively used to improve the operating range of centrifugal compressors.

Computational fluid dynamics (CFD), with advanced com-puting power, play a significant role in the analysis of flow field at near surge conditions in centrifugal compressors. Tamaki [1, 2] experimentally determined the cause of surge and improved the operating range by installing a ring groove

on the shroud casing. Sakaguchi [3] investigated the effect of a ring groove with and without guide vanes using numerical simulation to improve the surge margin of a centrifugal com-pressor for automotive turbochargers. Hembera [4] suggested the improvement of compressor performance by reducing the tip clearance, using the axial slot to remove the static leakage flow rate, and installing a groove to increase the intake flow.

A numerical investigation of the ring groove effect in a cen-trifugal compressor is presented. The flow fields at the leading edge of the impeller with and without the ring groove were analyzed in detail. Guide vanes were installed in the ring groove passage to improve the performance of the ring groove, and the consequent effects were studied. In addition, the com-pressor performance was compared by changing the location of the downstream slot.

2. Centrifugal compressor specifications and ring

groove charateristics

The compressor specifications are shown in Table 1. The meridional plane and the ring groove location are shown in Fig. 1. The upstream slot of the ring groove is located near the inducer, and the downstream slot is located between the lead-ing edge of the impeller and the splitter.

The flow directions in the ring groove are demonstrated in Fig. 2. The flow direction is forward at the near choking flow

*Corresponding author. Tel.: +82 41 589 8337, Fax.: +82 41 589 8330 E-mail address: [email protected]

† This paper was presented at the AJK2011, Hamamatsu, Japan, July 2011. Recom-mended by Guest Editor Hyon Kook Myong and Associate Editor Byeong Rog Shin.

© KSME & Springer 2012

1372 C.-Y. Park et al. / Journal of Mechanical Science and Technology 26 (5) (2012) 1371~1378

because the static pressure at the upstream slot is greater than the static pressure at the downstream slot. By contrast, the flow at the near surge has a reverse flow direction in the ring groove. Thus, the recirculation flow through the ring groove is combined with the main flow at the near surge flow rate. The inlet velocity triangles with and without the ring groove are shown in Fig. 3 at the near surge flow rate. The incidence angle is the deviation angle between the flow and blade angles. The relative velocity (W) is determined from the rotational speed (U) and the meridional velocity (Cm).

In general, the incidence angle is small at the near design flow rate, but the incidence angle increases as the flow rate decreases because of the reduced meridional velocity. If the incidence angle increases, then flow separation occurs and the shock loss increases at the leading edge of the impeller shroud. In addition, the stagnation leakage and stall flows are gener-ated near the shroud of the impeller. The unstable flow caused

by the stall and adverse pressure gradient are the important factors that results in surge formation. In the case of a ring groove system, some of the tip leakage flow are captured at the downstream slot and moved into the upstream slot. This recirculation flow merges with the incoming main flow at the impeller inlet. The reduced incidence angle caused by the increased meridional velocity and the absolute circumferential velocity component, which comes from the recirculation flow are shown in Fig. 3(b).

3. Numerical simulation

The impeller model was created using ANSYS-BladeGen and the structured grid was generated using ANSYS CFX- TurboGrid. Care must be undertaken in creating the narrow region between the impeller and the ring groove. Grid tests were performed from 300,000 hexahedral elements to 2.2 million hexahedral elements by checking the change in the recirculation mass flow rate through the ring groove. As a result, approximately 1.4 million hexahedral elements were selected. The grid topology showing some details of grid re-finement is shown in Fig. 4(a), and grid test results are shown in Fig. 4(b).

The compressor has vane diffusers downstream of the im-peller outlet. The total number of elements for the vane dif-fuser is approximately 80,000, whereas that for the ring groove passage is 100,000. An O-grid was applied around the impeller and diffuser wall. There are eight elements in the tip clearance region between the impeller and the shroud casing. Interfaces between the impeller and diffuser were coupled with the general grid interface (GGI) connection, and the stage average interface condition was used. Only one impeller pas-sage and two diffuser passages were analyzed with periodic conditions. As a boundary condition, total pressure and total temperature were considered to the inlet and mass flow rate or static pressure to the outlet as shown in Fig. 4(a).

ANSYS CFX-12, a commercial CFD code, was used for the numerical analysis. A three-dimensional RANS equation was used to analyze compressible turbulent flow. The shear stress

Table 1. Design specifications of centrifugal compressor.

Design mass flow rate (kg/s) 3.34

Total pressure ratio 4.8

Rotational speed (rpm) 41,500 Number of impeller blades

(main blade + splitter blade) 16

Fig. 1. Meridional view of ring groove.

Fig. 2. The recirculating flow in the ring groove.

(a) (b) Fig. 3. Velocity triangle at the impeller inlet near surge flow rate: (a) without; (b) with ring groove.

C.-Y. Park et al. / Journal of Mechanical Science and Technology 26 (5) (2012) 1371~1378 1373

transport k-ω model was used for the turbulent model. The flow was assumed to be steady. Determing the exact stall flow rate is challenging because the flow characteristics are very unstable at the near stall. However, the stall phenomenon could be related to the convergence characteristics in steady calculations. The predicted stall flow rate was set as the last stable flow rate to satisfy the following conditions. The rela-tive error between the inlet and outlet mass flow rates should be less than 0.5%, the total pressure ratio change should be less than 0.5% and the residual convergence should be less than 10-4. These conditions were determined by considering the convergence criterion of the axial compressor proposed by Huang [5] and Wang [6].

4. Numerical results and analysis

4.1 Comparison of numerical results with and without ring groove

The compressor performance, including the normalized pressure ratio and normalized efficiency, reflecting the exis-tence of the ring groove is illustrated in Fig. 5. The stall mar-gin, also an important performance parameter, is related to the

last stable flow and defined as follows:

(1 ) 100.s

design

QSMQ

= − × (1)

The stall margin improved by approximately 6%, with the

ring groove, but efficiency decreased by approximately 1% at the design point. We focused on the stall flow rate (Q/Qdesign =

(a)

0.14

0.15

0.16

0.17

0.18

0.19

0.2

0 5 10 15 20 25

Rec

ircul

atio

n flo

w (k

g/s)

Number of mesh (x10000) (b)

Fig. 4. (a) Boundary conditions; (b) Grid tests.

(a)

(b)

Fig. 5. Comparison of performance characteristics: (a) Total pressure ratio; (b) efficiency.

(a) (b) Fig. 6. Mach number contours at 90% span (Q/Qdesign = 0.883): (a) without; (b) with ring groove.


0.883) without the ring groove to determine the reason why the stall margin improved. The Mach number contours at 90% span are shown in Fig. 6. The flow through the impeller was transonic, and the shock loss, as well as the stall cell region, was reduced with the ring groove. Therefore, the flow was stable and the stall margin was improved.

The velocity components and flow angles near the impeller leading edge at the stall flow rate (Q/Qdesign = 0.883) without the ring groove were investigated to determine the flow char-acteristics of the ring groove. The meridional velocity distribu-

tion from hub to shroud is shown in Fig. 7(a). The meridional velocity increased above the 60% span (near shroud) when the ring groove was installed because of the recirculation flow through the ring groove passage. The jet flow at the upstream slot includes the circumferential velocity component, which is in the same direction as the impeller rotation. The circumfer-ential velocity increased rapidly at an over 80% span, as shown in Fig. 7(b). The incidence angle is represented by the difference between the blade and the flow angles. The flow angle is a combination of meridional velocity and the circum-ferential component, as shown in Fig. 3(b). The incidence angle distributions with and without the ring groove are shown in Fig. 7(c). When the ring groove was installed, the incidence angle was reduced from hub to shroud, especially at the near shroud. The reduced incidence angle suppressed flow separa-tion and made the flow more stable at the impeller leading edge.

4.2 Effect of guide vanes in the ring groove

The effect of guide vanes in the ring groove was examined by installing six guide vanes in the ring groove passage (Fig. 8). The flow fields with and without the guide vanes were

(a)

(b)

(c)

Fig. 7. Velocity and incidence angle distribution at impeller inlet(Q/Qdesign = 0.883): (a) Meridional velocity; (b) circumferential veloc-ity; (c) incidence angle.

Fig. 8. Ring groove geometry with and without guide vanes.

Fig. 9. Comparison of velocity vectors with and without guide vanes.


analyzed at the smallest converged flow rate (Q/Qdesign = 0.824). Without guide vanes, the flow direction through the ring groove passage was not uniform from the downstream to the upstream slot (Fig. 9(a)). There were large-scale back flows in the ring groove passage. These back flows prevented the incoming flow at the downstream of the ring groove. As a result, the recirculation flow was restricted. The recirculation flow was stabilized by installing the guide vanes, as shown in Fig. 9(b). Comparisons of compressor characteristics accord-ing to the presence of guide vanes are shown in Fig. 10.

The pressure ratio increased at low flow rates and efficiency improved as a result of installing the guide vanes. This im-provement can be explained by two key factors. First, the main flow rate increased because of the increased recirculation flow. As a result, flow separation was reduced by decreasing the incidence angle at the impeller leading edge. The normal-ized recirculation flow rates with and without guide vanes as the flow rate changes are shown in Fig. 11. The recirculation flow rate increased at low flow rates with guide vanes. Second, as the recirculation flow increased, the stall regions (low-velocity region) around the shroud were reduced by the cap-

ture of low-energy fluid in the stall region and transfer of this fluid into the downstream ring groove. This phenomenon also affects the impeller exit flow characteristics. The meridional velocity contours, especially at the near impeller exit and dif-fuser inlet, are shown in Fig. 12. With the guide vanes, the flow separations between the impeller exit and diffuser inlet near the shroud casing wall were reduced. The incidence angle distribution at the diffuser inlet is shown in Fig. 13. The large reverse zone near the shroud wall was reduced and stabilized by the guide vanes in the ring groove. As a result, the com-pressor performance improved. In addition, the stall margin increased by 2.8% because of the suppression of the flow sep-aration near the shroud casing of the impeller.

4.3 Effect of downstream ring groove location

Four different downstream locations along the shroud wall were considered to analyze the influence of the downstream

(a)

(b) Fig. 10. Comparison of performance characteristics: (a) Total pressureratio; (b) efficiency.

Fig. 11. Recirculation flow through the ring groove.

Fig. 12. Comparison of meridional velocity contours with and without guide vanes.


ring groove location (Fig. 14). The compressor performance curves based on the ring groove downstream location are illus-trated in Fig. 15. In the case of positions 2 and 3, the pressure ratio at low flow rates decreased compared with the original ring groove. When the ring groove downstream location changed to position 1, the pressure ratio decreased slightly, but the stall margin improved. These results are primarily caused by the flow direction in the ring groove and the velocity distri-bution near the shroud. As the location of the downstream slot moved toward the impeller exit, increased.

Therefore, the recirculation flow rate in the ring groove in-creased, and the inlet flow increased. The recirculation flow ratios for four ring groove locations are shown in Fig. 16. We determined that the amount of recirculation flow at position 2 is similar to that at position 3 at low flow rates regardless of the pressure differences between the upstream and down-stream slot. This phenomenon is caused by the back flow in the ring groove, as shown in Fig. 9. The back flow in the ring groove prevents the flow into the downstream slot, limiting the recirculation flows.

When the recirculation flow ratio is higher than 6%, the jet velocity in the upstream slot affects the main stream near the shroud wall. This condition is a very complex phenomenon. The total pressure contours and velocity vectors for positions 1 and 3 at low flow rates (Q/Qdesign = 0.824) illustrated in Fig. 17.

The interaction between the jet flow and main flow at posi-tion 1 was less than that at position 3 because the recirculation flow was relatively small.

Velocity contours at a 99% span for the four kinds of ring grooves at each stall flow rate are in Fig. 18. Of the various stall (low-velocity) areas in the plane in Fig. 18, the low veloc-ity region in the pressure side near the leading edge of the impeller (red circle in Fig. 18) is more effective in producing stall. The prescribed stall area increased as the slot location moved from position 1 to position 3. The incidence angle distribution at the minimum flow rate of the convergence cri-teria is shown if Fig. 19. The average incidence angle level at the stall flow rate decreased as the downstream slot position

Fig. 13. Incidence angle distribution at the diffuser inlet.

Fig. 14. Meridional location of downstream ring groove.

(a)

(b)

Fig. 15. Comparison of performance characteristics: (a) Total pressureratio; (b) efficiency.

Fig. 16. Recirculation flow ratio of ring grooves.


moved toward the impeller exit. Even though the average incidence angle level was high, as in the case of position 1, the effect of the small stall area in Fig. 18 was positive on the stall margin.

5. Conclusions

Numerical investigations were performed to examine the ring groove casing treatment to improve the operational range of a centrifugal compressor. The effect of ring groove casing treatment on the impeller inlet flow field was investigated near the stall flow rate.

The following conclusions have been drawn from this re-search.

(1) With the ring groove, the stall margin increased by 6%, but the efficiency decreased by 1% at the design flow rate.

(2) With the guide vanes in the ring groove system, the compressor performance improved at near the stall flow rate. The stall margin increased by 2.8% compared with the case without the guide vanes.

(3) The effect of the location of downstream slot on the stall margin and performance curve was considered. The incidence angle distribution and the low-velocity region in the pressure side near the leading edge of the impeller are important factors affecting the stall margin.

Nomenclature------------------------------------------------------------------------

Cm : Meridional velocity without ring groove Crm : Meridional velocity with ring groove Cθ : Velocity of circumferential SM : Stall margin PR : Pressure ratio Qs : Minimum flow rate of convergence criteria Qi : Inlet flow rate Qr : Recirculation flow rate U : Velocity of rotation W : Relative velocity without ring groove U : Relative velocity with ring groove

(a)

(b) Fig. 17. Total pressure and velocity vector at impeller inlet (Q/Qdesign = 0.824): (a) Position 1; (b) position 3.

Fig. 18. Comparison of velocity contours at 99% span.

Fig. 19. Incidence angle distribution at minimum flow rate of conver-gence criteria.


References

[1] H. Tamaki, H. Nakao and T. Aizawa, Experimental study on surge inception in a centrifugal compressor, The 7th AICFM, No. 50021 (2003).

[2] H. Tamaki, M. Unno and T. Kawakubo, Aerodynamic de-sign to increase pressure ratio of centrifugal compressors for turbochargers, ASME Turbo Expo, GT 2009-59160 (2009).

[3] D. Sakaguchi, K. Nagoshi, M. Tanimura, M. Ishida and H. Ueki, Effect of Guide vane in ring groove arrangement for a small turbocharger, The 10th AICFM, AICFM 0119 (2009).

[4] M. Hembera, H. P. Kau and E. Johann, Simulation of Casing Treatments of a Transonic Compressor Stage, International Journal of Rotating Machinery, ID 657202 (2008).

[5] X. Huang, H. Chen and S. Fu, CFD investigation on the circumferential grooves casing treatment of transonic com-pressor, ASME Turbo Expo, GT 2008-51107 (2008).

[6] T. Wang, W. Xu, C. Gu and C. Xiao, A new type of self-adaptive casing treatment for a centrifugal compressor, ASME Turbo Expo, GT 2010-23457 (2010).

Chi-Yong Park received his B.S. de-gree from Korea University of Technol-ogy and Education (KUT) in 2009, and his M.S degree from Hanyang Univer-sity, Korea, in 2011. He is currently a researcher in Advanced Engineering Team, Halla Climate Control Corp.

Young-Seok Choi received his B.S. degree from Seoul National University, Korea, in 1988, and his M.S. and Ph.D degrees in mechanical engineering at the same university in 1990 and 1996, re-spectively. He is currently a principal researcher in the Korea Institute of In-dustrial Technology (KITECH). His

research interests are CFD and design optimization of turbo-machinery.

Kyoung-Yong Lee received his B.S. and M.S. degree from Korea University of Technology and Education (KUT) in 2002 and 2004, respectively. He is con-currently in a course of Ph.D candidate in KUT and a researcher in KITECH. His research interests are system loss analysis and performance prediction of

turbomachinery.

Joon-Yong Yoon received his B.S. and M.S. degrees from Hanyang University and Ph.D degree from the Univ. of Iowa in mechanical engineering. He is a pro-fessor of mechanical engineering at Hanyang University. His research areas are CFD for applications, renewable energy, MEMS, and flow control.

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