Design of a Centrifugal Compressor with Low Specific Speed for Automotive Fuel Cell

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1 Proceedings of ASME Turbo Expo 2008: Power for Land, Sea and Air GT2008 June 9-3, 2008, Berlin, Germany GT Design of a Centrifugal Compressor with Low Specific Speed for Automotive Fuel Cell Xinqian Zheng, Yangjun zhang State Key Laboratory of Automotive Safety and Energy, Tsinghua University Beijing 00084, China Hong He National Key Lab. of Diesel Engine Turbocharging Tech. Datong, Shanxi , China Zhiling Qiu Department of Applied Mechanics,Chalmers University of Technology, Gothenburg, Sweden Abstract Centrifugal compressors driven by electric motor are the promising type for fuel cell pressurization system. A low specific speed centrifugal compressor powered by an ordinary high-speed (about 25,000rpm) electric motor has been designed at Tsinghua University for automotive fuel cell engines. The experimental results indicate that the designed low specific speed centrifugal compressor has comparatively high efficiency and wide operating range. In the condition of designed speed (24,000rpm), the highest efficiency and pressure ratio of the centrifugal compressor is up to 70% and.6, respectively. The designed low specific speed centrifugal compressor can meet the requirement of air systems of automotive fuel cell engines preliminarily. Moreover, the low specific speed centrifugal compressor avoids difficulties of usage of ultra-high-speed electric motors (about 60,000rpm) in high specific speed compressor. Based on the preliminary results of this centrifugal compressor, a new low specific speed centrifugal compressor with higher performances is being developed. NOMENCLATURE C absolute velocity L.E. M m blade leading edge meridional length mass flow m corrected mass flow cor n shaft speed n corrected shaft speed cor N s P p Q R T T.E. U W Z Δ h ω α β π η Subscripts non-dimensional specific speed power total pressure volumetric flow rate radius total temperature blade trailing edge rotational velocity relative velocity blade number specific isentropic enthalpy rotational angular speed absolute angle relative angle pressure ratio efficiency impeller inlet Copyright 2008 by ASME

2 2 impeller exit 5 diffuser exit 7 volute throat b h t blade hub tip INTRODUCTION Fuel cell vehicles (FCV) have captured the attention of more and more policymakers, environmentalists and automotive manufacturing companies due to their advantages of zero-emission, high efficiency and rapid response, etc. As the prime mover inside the FCV, the fuel cell system is pivotally important since it is directly related to vehicle performance. Fuel cell systems can use pressurizing air systems to improve performance. Pressurized fuel cell systems have higher power density, higher system efficiency, and better water balance than atmospheric fuel cell systems [, 2]. The requirements of pressurized fuel cell systems include: compact structure, light weight, oil-free of air, low operate noise, easy maintenance, low cost, and high operation efficiency [3]. Nowadays, displacement compressors are widely used in air pressurizing systems of fuel cells due to their high pressure ratio with low shaft speed and low mass flow, among which screw compressors are fairly representative. However, screw compressors are not able to operate with turbines recovering exhaust gas power because of low shaft speed limitation; therefore they cannot manage to save energy as centrifugal compressors combining turbines. Moreover, centrifugal compressors have more advantages of quick response, long longevity, and high efficiency, etc. So, centrifugal compressors are considered as one of the most prospective pressurization systems. However, nowadays it is necessary to use some auxiliary power to operate centrifugal compressors since exhaust gas power of fuel cell systems till now is not powerful enough to drive turbines working with centrifugal compressors. As one of the auxiliary powers, an ultra-high-speed electric motor (about 60,000 rpm) can be used to drive centrifugal compressors. A turbocharger type compressor system integrated with an ultrahigh-speed (>00,000 rpm) motor is being developed by Honeywell [4]. This system makes fuel cell pressurizing air system more compact and has quicker response characteristic. However, usage of ultra-high-speed motor causes several serious problems including high cost, complex maintenance, low stability, need of special cooling system, etc. Therefore this type of compressor system is hard to be applied in commercialization products [5]. Using ordinary high-speed motors (about 25,000rpm) can avoid above problems of ultra-high-speed motors. However, if centrifugal compressors are still designed conventionally in this case, pressure ratio can only achieve around., therefore cannot meet the pressurizing requirements of air systems for fuel cells. So it is necessary to design a centrifugal compressor with higher pressure ratio in the same condition of rotational speed and mass flow, i.e., to design a low specific speed centrifugal compressor. However, in low specific speed condition, it is a challenge to design a centrifugal compressor with high efficiency and wide operation range characteristic Moreover, there is few reference relating to this type of design. Compressor design software, Concept NREC, has been used to design and analyze the low specific speed centrifugal compressor [6]. The experimental results show that designed low specific speed centrifugal compressor has a reasonable performance. 2 CENTRIFUGAL COMPRESSOR DESIGN The main design parameters of the low specific speed compressor are: design rotational speed 24,000 rpm, design pressure ratio.6, design flow 0.5 kg/s. 2. Low Specific Speed Design Non-dimensional specific speed N s is given by: /2 Q Ns = ω ( ω is the rotational speed, Q is the Δ 3/4 h volumetric flow rate, Δ h is the specific isentropic enthalpy). Specific speed qualitatively shows work ability of compressors. With same rotational speed, lower specific speed means higher compressor ratio. Specific speed for conventional design of centrifugal compressors is typically around between 0.7 and.0 so that compressors can achieve comparatively high efficiency. However, compared to conventional design, low specific speed centrifugal compressors have higher inverse pressure gradient and larger fraction of secondary flow. Therefore profound understanding of flow characteristic inside stage is needed to design a low specific speed centrifugal compressor with high performance. 2.2 One-dimensional Design Figure shows the meridional view of the low specific speed centrifugal compressor, which can be specified by the following charactistics. A) Long flow passage design: Inducer of impeller uses comparatively small radius of curvature and adopts zero-incidence design to be suitable for low flow rate work condition of fuel cell systems. Elongate blade design is required to meet the demand of high pressure ratio within low rotational speed and low flow rate. Moderately big impeller radius is needed to generate high tangent velocity of exit, therefore, make compressors able to output high pressure ratio under low rotational speed. But the maximum impeller radius is limited by the requirement of compact design of fuel cell systems. B) Short vaneless-diffuser design: Diffuser is used to converts kinetic energy of air exiting from impeller into static pressure. Automotive centrifugal compressor normally uses vaneless diffuser due to wide operation range and small volume. Air flows almost as a logarithmic curve inside of vaneless 2 Copyright 2008 by ASME

3 diffuser. At the exit of low specific speed impeller, tangential velocity of air is much higher than radial velocity of air along impeller blade, which indicates that absolute velocity of airflow inside vaneless diffuser is quite close to tangential direction (see Fig.2). As a result, the distance air flows in the vaneless diffuser greatly increases so that frictional losses are high and therefore back flow occurs easily. In order to reduce these losses, short vaneless diffuser is required in this case. Inlet total pressure P = Inlet total temperature T = 00 kpa 298 K Design mass flow rate m = 0.5 kg/s Design shaft speed n= 24,000 rpm Design specific speed N s = Blade count full/splitter Z f /Z s = 0/0 Tip radius at impeller inlet R t = 3 mm Hub radius at impeller inlet R h = 7 mm Blade angle at L.E. tip β tb = deg Exit blade radius R 2 = 08 mm Exit blade depth BB2= 5 mm Exit blade angle β 2b = -40 deg Blade rotational velocity at exit U 2 27 m/s Average exit radius of vaneless diffuser R 5 = 0 mm volute Throat hydraulic diameter of D 7 = 30 mm Fig. Meridional view of centrifugal compressor design Fig. 2 Velocity triangle at blade exit The main parameters of the designed centrifugal compressor are presented in table. Compal module of Concepts NREC is used to optimize the parameters, such as the effects on the performance from impeller's exit diameter, blade angel, exit width and so on. The method of performance prediction is described in Ref. [6]. Table : Main parameters of centrifugal compressor 2.3 Impeller Three-dimensional Design Axcent module of Concept NREC is used to design the impeller. The impeller blade is designed to apply loading distributions along hub and shroud meridional traces. A linear connection between points of both meridional traces along quasi-orthogonal lines generates a ruled surface which make it possible to manufacture the blades by flank milling. The shape of a typical blade is defined by means of curves that specify the blade angle distribution and blade thickness distribution. Of course, the hub and shroud contours should be defined too. Blade angle/thickness distribution is defined by using a Bezier curve. The Bezier method defines the curve segment in terms of a polygon, two of whose vertices are the end points of the segment. The intermediate polygon points will not, in general, lie on the curve. On a Bezier curve, the effect of moving a point varies with the proximity to that point. The shape of the curve is most affected in the region close to the point that is moved. Smaller changes occur farther away; however, the whole curve is changed to some extent, although changes far from the point moved are usually quite negligible. In order to define the blade geometry, an initial assessment of the quality of the design could be obtained by means of MST (multi-streamtube) approach. The MST analysis is a pure streamline curvature technique that solves a velocity gradient equation along quasi-orthogonals, used to determine the velocity distribution from hub to shroud and linearized blade to blade. MST method is known to be comparatively stable, fast, and unique in its resulting calculations. By performing a MST 3 Copyright 2008 by ASME

4 calculation before any CFD (Computational Fluid Dynamics) analysis, the design could be more productive. MST is first used to configure the blade shape to obtain a design that comes within recommended loading limits for the blades. Then, CFD is used to confirm that there are no regions of separated or reversed flow. If there are regions of adverse flow conditions, we can make a design change, revise the blade shapes in MST, and then confirm the design improvement in CFD. The detailed design process is described in Ref. [6]. Blade angle distribution is very important for loading distribution in terms of the Euler turbomachine equation. Figure 3 shows the blade angle distribution along hub and shroud. Abscissa is the dimensionless meridional trace length %M = [M(i)-M(0)/[M(T.E.)-M(0)]*00%; M(0) is the meridional L.E. position of both traces. The L.E. position of splitter blade is shown in Fig.. The splitter blade angle distribution follows the main blade. Figure 4 shows the wrap angle (theta distribution). If the wrap starts at zero, this shows how far around the axis of rotation the blade is wrapped. Figure 5 displays a plot of the blade lean angel. This is the angle that the curve represents the difference between the hub and shroud wrapping; the further apart the hub and the shroud wraps become, the more the blade leans. The lean also correlates to other aspects of the design, including the radius. Due to these correlations, the blade angle, wrap angle, and lean angle are all related. Fig. 5 Blade wrap angle distribution Fig. 6 Blade lean angle distribution Fig. 3 Blade angle distribution 4 EXPERIMENTS After optimization design, the centrifugal compressor is manufactured. The impeller photo is shown in Fig. 7. Impeller are made of aluminum alloy and machined by five-axis NC milling machine. In order to obtain the experimental performance conveniently, the designed centrifugal compressor is driven by a turbine just like an experiment for a turbocharger. The experiment rig photo is shown in Fig. 8. The turbine is driven by the pressured air which is hot up in combustion. The total pressure p is measured by total pressure rake with ± 0.2% inaccuracy. Total temperature T is measured by thermocouple with inaccuracy ± 0.5. Mass flow m is measured by vortex flowmeter with inaccuracy ±%. 4 Copyright 2008 by ASME

At designed speed of 24,000rpm, the highest efficiency and pressure ratio of the centrifugal compressor is up to 70% and.6, respectively.

5 298K ncor = n (3) Tt Then, the power contours in Fig. 0 are calculated by the following formula P = m cor Cp ( Tt 2 T t) (4) Figure 9 shows pressure ratio and efficiency contours of the centrifugal compressor with low specific speed. At designed speed of 24,000rpm, the highest efficiency and pressure ratio of the centrifugal compressor is up to 70% and.6, respectively. The maximum efficiency ring is 78%, which is close to the efficiency of conventional centrifugal compressor. Figure 0 shows pressure ratio and power contours of the centrifugal compressor. Power consumption of compressor is important for performance of full cell. At the designed point ( m cor =0.5kg/s, n cor =24,000rpm), Power consumption of compressor is just about 8.5kW..90 Fig. 7 Impeller photo Fig. 8 Experimental rig photo The experimental results including pressure ratio, efficiency contours, power contours, corrected mass flow and corrected rotation speed are plotted in Fig. 9 and Fig. 0. The pressure ratio π is the ratio of absolute outlet pressure P divided by absolute inlet pressure. The efficiency 2 P contours in Fig. 9 is defined as k k ( π ) η = () Tt2 Tt The corrected mass flow is defined as Tt 00kPa m cor = m (2) 298K Pt The corrected shaft speed is defined as Pressure ratio % 78% 75% 4000rpm 2000rpm 8000rpm0000rpm 8000rpm 6000rpm 28000rpm 26000rpm 24000rpm 22000rpm 20000rpm Corrected mass flow (kg/s) Fig. 9 Experimental performance map of centrifugal compressor: pressure ratio & efficiency contours. 68% 65% 60% 5 Copyright 2008 by ASME

6 Pressure ratio kw 2kW 4kW 7kW 0kW 4000rpm 2000rpm 8000rpm 0000rpm 5kW 20kW 28000rpm 26000rpm 24000rpm 22000rpm 20000rpm 8000rpm 6000rpm Corrected mass flow (kg/s) Fig. 0 Experimental performance map of centrifugal compressor: pressure ratio & power contours. [2] Galen, W. K., Myron, A. H., 200, A Comparison of Two Air Compressors for PEM Fuel Cell Systems, Virginia: Virginia Polytechnic Institute and State University. [3] Vine, A. J., Thornton, W. E., Pullen, K. R., et al, 2005, Low Specific Speed Turbocompressors, International Conference on Compressors and their Systems, London: Springer: [4] Mark, G., 2004, Cost and Performance Enhancements for a PEM Fuel Cell Turbocompressor, DOE Hydrogen, Fuel Cells & Infrastructure Technologies Program Review Presentation, Honeywell Systems, Systems & Services, Philadelphia. [5] Chen, Q. S., Qi, Z. N., 200, Technology Challenge and Prospect of Fuel Cell Vehicle, Automotive Engineering, 23(6): (in Chinese) [6] Japikse, D., 2006, Centrifugal Compressor Design and Performance, Vermont: Concepts ETI, Inc., Wilder. 4 CONCLUSIONS A low specific speed centrifugal compressor for automotive fuel cell systems is designed in this project. This centrifugal compressor can be driven by an ordinary high-speed electric motor (about 25,000rpm), therefore has advantages of low cost, easy manufacture, long longevity to be fit for low flow rate condition. The experimental results show that designed low specific speed centrifugal compressor reaches efficiency of 70 % and pressure ratio of.6 at the design rotational speed of 24,000 rpm, which basically fulfils the pressurizing requirements of fuel cell systems. The maximum efficiency ring is up to 78%, which is close to the efficiency of conventional centrifugal compressor. Based on the preliminary results of this centrifugal compressor, a new low specific speed centrifugal compressor with higher performances is being developed. 5 ACKNOWLEDGE This project has been supported by Xuyao Chen, Fenghu Liu in FuYuan Turbochargers CO., LTD (Weifang, Shandong, China). REFERENCES [] Cunningham, 200, A Comparison of High-Pressure and Low-Pressure Operation of PEM Fuel Cell Systems, SAE 200 World Congress, Detroit, Copyright 2008 by ASME

Alpha College of Engineering

Alpha College of Engineering Department of Mechanical Engineering TURBO MACHINE (10ME56) QUESTION BANK PART-A UNIT-1 1. Define a turbomahcine. Write a schematic diagram showing principal parts of a turbo