2008_Design of a Centrifugal Compressor with Low Specific Speed for Automotive Fuel Cell

g centrifug designed the requ prelimin compres electric compres compres with hig NOMEN C L.E. M m� corm� n Proceedings of ASME Turbo Expo 2008: Power for Land, Sea and Air GT2008 June 9-13, 2008, Berlin, Germany ental results indicate that the designed low specific ntrifugal compressor has comparatively high efficiency e operating range. In the condition of designed speed pm), the highest efficiency and pressure ratio of the al compressor is up to 70% and 1.6, respectively. The low specific speed centrifugal compressor can meet irement of air systems of automotive fuel cell engines arily. Moreover, the low specific speed centrifugal sor avoids difficulties of usage of ultra-high-speed motors (about 60,000rpm) in high specific speed sor. Based on the preliminary results of this centrifugal sor, a new low specific speed centrifugal compressor her performances is being developed. CLATURE absolute velocity blade leading edge meridional length mass flow corrected mass flow shaft speed Q volumetric flow rate R radius T ∗ total temperature T.E. blade trailing edge U rotational velocity W relative velocity Z blade number hΔ specific isentropic enthalpy ω rotational angular speed α absolute angle β relative angle π pressure ratio η efficiency Subscripts 1 impeller inlet 1 Copyright © 2008 by ASME Design of a Centrifugal Comp Automotive Xinqian Zheng, Yangjun zhang State Key Laboratory of Automotive Safety and Energy, Tsinghua University Beijing 100084, China Zhilin Department Mechanics ,Chalm Technology, Goth 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 experim speed ce and wid (24,000r ressor with Low Specific Speed for Fuel Cell Hong He National Key Lab. of Diesel Engine Turbocharging Tech. Datong, Shanxi 037036, China Qiu of Applied ers University of enburg, Sweden corn corrected shaft speed sN non-dimensional specific speed P power p∗ total pressure GT2008-50468 2 Copyright © 2008 by ASME 2 impeller exit 5 diffuser exit 7 volute throat b blade h hub t tip 1 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 [1, 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 ultra- high-speed (>100,000 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 1.1, 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 1.6, design flow 0.15 kg/s. 2.1 Low Specific Speed Design Non-dimensional specific speed Ns is given by: 1/ 2 3/ 4s QN h ω= ⋅ Δ ( ω is the rotational speed, Q is the volumetric flow rate, hΔ 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 1.0 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 1 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 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. Fig. 1 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 1. 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 1: Main parameters of centrifugal compressor Inlet total pressure 1P ∗ = 100 kPa Inlet total temperature 1T ∗ = 298 K Design mass flow rate m� = 0.15 kg/s Design shaft speed n= 24,000 rpm Design specific speed Ns= 0.265 Blade count full/splitter Zf/Zs= 10/10 Tip radius at impeller inlet R1t= 31 mm Hub radius at impeller inlet R1h= 17 mm Blade angle at L.E. tip β1tb= -53.4 deg Exit blade radius R2= 108 mm Exit blade depth BB2= 5 mm Exit blade angle β2b= -40 deg Blade rotational velocity at exit U2 271 m/s Average exit radius of vaneless diffuser R5= 110 mm Throat hydraulic diameter of volute D7= 30 mm 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 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)]*100%; M(0) is the meridional L.E. position of both traces. The L.E. position of splitter blade is shown in Fig. 1. 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. 3 Blade angle distribution Fig. 5 Blade wrap angle distribution Fig. 6 Blade lean 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 ±1%. � 4 Copyright © 2008 by ASME 5 Copyright © 2008 by ASME 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. 10. The pressure ratio π is the ratio of absolute outlet pressure divided by absolute inlet pressure . The efficiency contours in Fig. 9 is defined as 2P ∗ 1P ∗ 1 2 1 ( ) 1 1 k k t tT T πη − −= − (1) The corrected mass flow is defined as 1 1 100 298 t cor t T kPam m K P = × ×� � (2) The corrected shaft speed is defined as T f t s t T e s c p a F c 1 298 cor t Kn n T = × (3) hen, the power contours in Fig. 10 are calculated by the ollowing formula 2 1(cor p t tP m C T T )= × × −� (4) Figure 9 shows pressure ratio and efficiency contours of he centrifugal compressor with low specific speed. At designed peed of 24,000rpm, the highest efficiency and pressure ratio of he centrifugal compressor is up to 70% and 1.6, respectively. he maximum efficiency ring is 78%, which is close to the fficiency of conventional centrifugal compressor. Figure 10 hows pressure ratio and power contours of the centrifugal ompressor. Power consumption of compressor is important for erformance of full cell. At the designed point ( =0.15kg/s, =24,000rpm), Power consumption of compressor is just bout 8.5kW. corm� corn 1.00 1.10 1.20 1.30 1.40 1.50 1.60 1.70 1.80 1.90 0.00 0.05 0.10 0.15 0.20 0.25 0.30 0.35 Corrected mass flow (kg/s) Pr es su re ra tio 14000rpm 16000rpm 18000rpm 20000rpm 22000rpm 28000rpm60% 65% 68% 78% 12000rpm10000rpm8000rpm 75% 72% 26000rpm 24000rpm ig. 9 Experimental performance map of centrifugal ompressor: pressure ratio & efficiency contours. W. K., Myron, A. H., 2001, “A Comparison of Two mpressors for PEM Fuel Cell Systems,” Virginia: a Polytechnic Institute and State University. . J., Thornton, W. E., Pullen, K. R., et al, 2005, Specific Speed Turbocompressors,” International ence on Compressors and their Systems, London: er：225-234. G., 2004, “Cost and Performance Enhancements for a uel Cell Turbocompressor,” DOE Hydrogen, Fuel & Infrastructure Technologies Program Review tation, Honeywell Systems, Systems & Services, lphia. Q. S., Qi, Z. N., 2001, “Technology Challenge and ct of Fuel Cell Vehicle,” Automotive Engineering, 361-364. (in Chinese) , D., 2006, “Centrifugal Compressor Design and ance,” Vermont: Concepts ETI, Inc., Wilder. Copyright © 2008 by ASME 1.00 1.10 1.20 1.30 1.40 1.50 1.60 1.70 1.80 1.90 0.00 0.05 0.10 0.15 0.20 0.25 0.30 0.35 Corrected mass flow (kg/s) Pr es su re ra tio 20kW 15kW 10kW 7kW 1kW 4kW 2kW 8000rpm 10000rpm12000rpm 14000rpm 16000rpm 18000rpm 20000rpm 22000rpm 24000rpm 26000rpm 28000rpm Fig. 10 Experimental performance map of centrifugal compressor: pressure ratio & power contours. 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 1.6 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 [1] Cunningham, 2001, “A Comparison of High-Pressure and Low-Pressure Operation of PEM Fuel Cell Systems,” SAE 2001 World Congress, Detroit, 2001-01-0538. [2] Galen, Air Co Virgini [3] Vine, A “Low Confer Spring [4] Mark, PEM F Cells Presen Philade [5] Chen, Prospe 23(6): [6] Japikse Perform 6