Mihăescu, M

Semlitsch, B

Sundström, E

English

Elias Sundström

Bernhard Semlitsch

Mihai Mihăescu

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Similarities and Differences Concerning Flow Characteristics in Centrifugal Compressors of Different Size

Apollo (Cambridge)

Similarities and differences concerning flowcharacteristics in centrifugal compressors ofdifferent sizeElias Sundstro¨m, Bernhard Semlitsch and Mihai Miha˘escu1 IntroductionDownsizing of an automotive internal combustion engine is playing an importantrole for improving the energetic efficiency and reducing pollutant emissions [1].Turbocharging is a key and matured technology, used for achieving these aims.However, its application range is bound to certain limits by appearance of unwantedfluid-flow phenomena leading to conditions like surge at low mass-flow rates andchoke at high mass-flow rates [2]. The lack of detailed knowledge concerning thecompressible flow behavior associated with turbo-machinery at off-design operatingconditions imposes restrictive security margins on turbocharger’s operating range.This is in order to avoid the occurrence of the violent flow phenomena, as e.g.surge, which can damage the compressor. Nevertheless, the surge line (indicatingwhen surge occurs) is experienced to be individual for each compressor design anda generalization of surge limit to other designs is a desired but challenging task.Commonly, the surge phenomenon limits the application range more at the highrotational speed lines than at low rotational speed lines.Gas stand measurements are carried out for experimental assessment of tur-bochargers, to quantify their performance parameters [3]. Such tests cannot givean indication about the three-dimensional flow and pressure fields associated withthe different operating conditions. Nevertheless, the examination of the flow field isElias Sundstro¨mLinne´ FLOW Centre and CCGEx, Department of Mechanics, KTH Royal Institute of Technology,Osquars Backe 18, 10044 Stockholm, Sweden, e-mail: elias@mech.kth.seBernhard SemlitschLinne´ FLOW Centre and CCGEx, Department of Mechanics, KTH Royal Institute of Technology,Osquars Backe 18, 10044 Stockholm, Sweden, e-mail: bernhard@mech.kth.seMihai Miha˘escuLinne´ FLOW Centre and CCGEx, Department of Mechanics, KTH Royal Institute of Technology,Osquars Backe 18, 10044 Stockholm, Sweden, e-mail: mihai@mech.kth.se12 Elias Sundstro¨m, Bernhard Semlitsch and Mihai Miha˘escurelevant for understanding the behavior of the centrifugal compressor under variousoperating conditions.In the present analysis, steady-state, three-dimensional Reynolds Averaged Navier-Stokes (RANS) simulations are used to explain the differences in the compressorperformance maps for two automotive centrifugal compressors of different size.Special emphasis is put on the conditions closer to the surge line, over differentspeed lines.2 Geometry and Numerical MethodThe two investigated compressor geometries and the corresponding specificationsare shown in Fig. 1. One is named small compressor (diffuser area ratio = 0.56)and the other is named large compressor (diffuser area ratio = 0.62). The surge by-pass channel with a lock valve, the ball bearing features, and the impeller nut havebeen removed for simplification of the simulation setup. The inlet and volute exitpipes are extruded several diameters in the boundary normal directions, where thepiping systems are equal for both compressors. Hence, an idealized installation ofthe compressor is considered (replicating the gas stand experiment) under so-calledcold conditions with adiabatic wall boundary conditions.In order to characterize the flow behavior occurring in the two compressors,a large number of operating conditions need to be simulated (different rotationalspeeds and mass-flow rates). Computational techniques based on the steady-stateRANS formulation are relatively fast, robust, and affordable as compared with un-steady approaches [4], which can capture the mean flow features and trends corre-sponding to the measured global performance parameters. Despite their limitations,steady-state RANS modeling can be utilized for covering an impressive number ofFig. 1 Side and front views of the compressor geometry, small compressor at the top and the largecompressor at the bottom. The mesh resolution is indicated in the enlarged impeller surface insets.Flow characteristics in centrifugal compressors of different size 3cases within the stable compressor operating conditions, in a relatively short com-putational time [5].Industry-standard modeling is employed to provide fast (3 hour) turn-around ateach compressor operating point on a 6-core CPU workstation. The flow govern-ing equations are solved by a compressible implicit coupled solver [6] based on thefinite volume approach. The two-equation k−ω SST turbulence model is applied.About one million polyhedral computational cells are used to discretize the com-putational domain of each of the compressors, where two prism-layers are placedtowards the no-slip walls. The Moving Reference Frame (MRF) approach is consid-ered to take the effects of impeller’s rotation into account. Mixing-plane interfacesare used which gives a circumferential averaged flow at the interfaces between theimpeller region and the stationary. This treatment allows for mean radial variations.3 ResultsThe global performance prediction in terms of the total-to-total pressure ratio andtotal-to-total efficiency are compared in Fig. 2 for the two compressor geometries.An overall good agreement between the numerical predictions and the experimentalresults could be achieved. A more pronounced humpback feature at the peak pres-sure ratios for the higher speed-lines is estimated by the computations. The geomet-rical simplifications for the numerical setup, as e.g. the removal of the surge by-passchannel, could potentially allow that higher pressure ratios are achieved at the higherspeed-lines. The predicted total-to-total efficiency is in an overall good agreementsmallcompressor0 0.5 11.52.53.5Normalized corrected mass flow rate (−)PRt−t (−)  ExperimentSimulationnorm uredC =23222019181513100 0.5 10.50.60.7Normalized corrected mass flow rate (−)Efficiency t−t (−)2322201918151310largecompressor0 0.5 11.52.53.5Normalized corrected mass flow rate (−)PRt−t (−)norm uredC =23222019181513100 0.5 10.50.60.7Normalized corrected mass flow rate (−)Efficiency t−t (−) 2322201918151310Fig. 2 Compressor pressure ratio (PR) and efficiency vs. normalized corrected mass flow.4 Elias Sundstro¨m, Bernhard Semlitsch and Mihai Miha˘escuwith the experimental data although the values are generally over-predicted by thesimulation, where an improvement can be seen with higher mass flow rates. Theover-prediction in absolute levels can be related to the not identical boundary condi-tions between the computational and experimental setup. Adiabatic wall treatmentis employed in the simulation setup, whereas the turbocharger is running under so-called hot conditions in the experiment. Therefore the omission of wall heat transferis seen to give an over-prediction at low mass flow rates, but at higher mass flowrates the assumption of an adiabatic wall becomes less crucial [7].The flow field characteristics of the two investigated centrifugal compressors aregiven in Fig. 3 for the same impeller speed-line (normalized uredC = 19), which havebeen designed to operate in different mass-flow rate regimes. The smaller compres-sor is aimed for the usage at lower mass-flow rates, while the larger compressorcan cope with larger mass-flow rates for e.g. a higher performance engine. The flowenters from the left into the domain and is accelerated towards the impeller. Thecompressor wheel is turning into the clockwise direction. At the tips of the wheel,the highest flow velocities are provoked. The imposed kinetic energy by the wheelblades is converted into pressure energy in the diffuser (the flow velocities decrease).From the diffuser, the stream exits tangentially into the volute. Since the volute in-creases its cross-sectional volume in clockwise direction (consistent with the wheelrotation), the pressure increases towards the outlet, see Fig. 3 (b). The observed flowfeatures for the two compressor designs are similar at the near optimum efficiencyoperating condition.Towards the surge line, lower mass-flow rates push against a high outlet pressure.The generated kinetic flow energy is not high enough to overcome the high staticpressure built at the outlet and local flow reversal occurs. Thus, the flow is pushedfrom the diffuser region through the tip leakage (at the periphery of the impeller)against the main flow direction [8]. This is obvious for both compressors in Fig. 3(a), where the annular high velocity flow can be seen at the outer periphery. This(a) Last stable operating point (b) Near optimum efficiencyFig. 3 Mach number and pressure coefficient contours in side and front view planes (small com-pressor at the top rows and large compressor at the bottom rows).Flow characteristics in centrifugal compressors of different size 5flow feature is characteristic towards the surge margin and lowers the compressorperformance. The reversed flow emerging from the tip leakage reaches far furtherupstream for the larger compressor. This is an effect of the larger compressor oper-ating at higher mass-flow rates towards its surge line. The interaction between thereversed flow near the periphery of the impeller and the incoming flow triggers thesurge occurrence at different locations in the compressor map.At this near surge operating point, several differences in the flow field can be ob-served. The smaller compressor is capable to operate at lower mass flow rates, whichresults in lower flow velocities at the impeller inlet (see Fig. 3 (a)). This affects alsothe recirculating tip leakage flow, which is larger for the larger compressor. An-other difference is the shift of the low-pressure region to the tongue region, which isslightly more pronounced for the small compressor. The compressor volute exhibitsa significantly different design, which generates a different flow field appearance.For the large compressor, a design is chose allowing the flow to be attached at theouter volute walls. Generally, the same swirling motion as for the near optimum effi-ciency operating condition is induced in the volute, but a small stagnant flow regionis provoked due to the low flow momentum. In contrary, the diffuser flow detachesfrom the walls and leaks centrally into the volute for the small compressor.Figure 4 shows that the velocity upstream of the blades near the shroud increaseswith higher speed-lines for the last stable operating point, since the mass flow rategenerally increases along the last stable point in the compressor map. The mass flowrate is lower near the surge-line for the lower speed-lines, which induces reducedvelocities and the blades operate in an overall subsonic flow regime. In contrary,local sonic bubbles manifest at the blade tips for higher speed-lines, where the flowis accelerating over the suction side of the blades. Separated flow can be observed onthe suction side of both the main and splitter blades. It commences on the mid-bladespan of the main blades and extends to the rear of the blades. Towards the hub, thesmallcompressor19 20 22 23largecompressor19 20 22 23Fig. 4 Velocity vectors colored by the Mach number on a flattened surface at 85% blade span forthe last stable operating point (increasing speed-lines uredC = 19 to uuredC = 23 from left to right).6 Elias Sundstro¨m, Bernhard Semlitsch and Mihai Miha˘escuflow remains more attached at the blades, whereas close to the shroud occurs the tipleakage for these operating conditions. This flow separation decreases the impellerefficiency. By the diffuser inlet, the flow accelerates expanding in the volute.4 ConclusionsThe global compressor performance in terms of the pressure ratio and the efficiencyare predicted numerically using a steady-state RANS approach for two differentcompressor geometries. A good agreement with measurements for an idealized in-stallation are achieved for a vast range of operating conditions at different speed-lines. The observed flow features for the two compressor designs are similar at thenear optimum efficiency operating condition. However, towards the near surge op-erating condition, the different compressor design effects the flow field. The largerinducer diameter provokes lower velocities for a equal mass flow rates, which resultsin a less stable operating performance when approaching the surge line.Acknowledgements The Competence Center for Gas Exchange (CCGEx), the Swedish EnergyAgency (STEM), Volvo Group and SCANIA are acknowledged for supporting this project. Thecomputations were performed on resources provided by the Swedish National Infrastructure forComputing (SNIC) at PDC Centre for High Performance Computing (PDC-HPC) and High Per-formance Computing Center North (HPC2N). Special thanks to BorgWarner for sharing the CADgeometries.References1. W. Knecht. Diesel engine development in view of reduced emission standards. Energy, 33(2),264-271, 2008.2. B. Semlitsch, V. JyothishKumar, M. Miha˘escu, L. Fuchs, E. Gutmark. Investigation of theSurge Phenomena in a Centrifugal Compressor Using Large Eddy Simulation. ASME 2013IMECE, IMECE2013-66301, 2013.3. J. Galindo, F. Arnau, A. Tiseira, R. Lang, H. Lahjaily, T. Gimenes. Measurement and Mod-eling of Compressor Surge on Engine Test Bench for Different Intake Line Configurations.SAE Technical Paper, 2011-01-0370, 2013.4. O. Baris, F. Mendonca. Automotive Turbocharger Compressor CFD and extension towardsincorporating installation effects. ASME Turbo Expo, GT2011-46796, 2011.5. E. Sundstro¨m, B. Semlitsch, M. Miha˘escu. Assessment of the 3D flow in a centrifugal com-pressor using steady-state and unsteady flow solvers. SAE Technical Paper, 2014-01-2856,2014.6. STAR-CCM+, release version 10.02, CD-adapco, February 2015, www.cd-adapco.com.7. J. Serrano, P. Olmeda, F. Arnau, M Reyes-Belmonte, A. Lefebvre. Importance of heat transferphenomena in small turbochargers for passenger car applications. SAE Int. J. Eng. 6(2), 716-728, 2013.8. B. Semlitsch, V. JyothishKumar, M. Miha˘escu, L. Fuchs, E. Gutmark, M. Gancedo. Numer-ical flow analysis of a centrifugal compressor with ported and without ported shroud. SAETechnical Paper, 2014-01-1655, 2014.

Similarities and differences concerning flow characteristics in centrifugal compressors of different size

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Similarities and differences concerning flow characteristics in centrifugal compressors of different size

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