J. of Thermal Science Vol.12, No.1
Design and Prototyping of Micro Centrifugal Compressor Shimpei M i z u k i 1
Gaku Minorikawa 1
Naoki Y a m a g u c h i 1
Yutaka O h t a z
Toshiyuki Hirano I
Yuichiro Asaga l
Eisuke O u t a 2
1. D e p a r t m e n t of Mechanical Engineering, Hosei University, 7-2, Kajinocho 3 chome, Koganei-shi, Tokyo, 184-8584, Japan 2. Department of Mechanical Engineering, Waseda University, 4-1, O h k u b o 3 chome, Shinjuku-ku, Tokyo, 169-8555, Japan
In order to establish the design methodology of ultra micro centrifugal compressor, which is the most important component of ultra micro gas turbine unit, a 10 times of the final target size model was designed, prototyped and tested. The problems to be solved for downsizing were examined and 2-dimensional impeller was chosen as the first model due to its productivity. The conventional 1D prediction method, CFD and the inverse design were attempted. The prototyped compressor was driven by using a turbocharger and the performance characteristics were measured.
Keywords: centrifugal compressor, performance characteristics, 2-dimensional impeller.
Introduction Centrifugal compressors are widely used in various fields such as industry, aviation and vehicle. The amount of products has been increasing dynamically since small compressors were used as turbochargers for automobile engines. Thus, it seems that the methodology of design and the production techniques have been matured. On the other hand, many researches for micro gas turbines for separated power generations have been tried in recent years. However, it is still unclear for the design methodology of ultra micro gas turbines, which is much smaller than conventional micro gas turbines using for mobile power units, micro jet engines and so on. The present study is an attempt to establish the design method of an ultra micro centrifugal compressor, which is the most important component of a ultra micro gas turbine system. In this study, a 10 times size of the final target compressor was designed, prototyped and tested.
Design of Ultra-micro Centrifugal Compressor Problems to be solved A few studies of ultra-micro compressors had been Received 2002
reported, for the impellers with 2 or 3 dimensional configurationsfl ~41. It is well known that the performance characteristics of impellers are getting better as the configurations are changed from 2 dimensional to quasi 3 dimensional and fully 3 dimensional tSl. On the other hand, the difficulty for the manufacturing is increasing. In this study, as a first step of ultra small compressors, the ten times size of the final target was examined. By using the conventional machining, it is possible to make 3-dimensional configurations, even if the size is getting a little smaller. However, taking the productivity of the target size into account, the 2-dimensional shape was chosen in this study. When the Reynolds number based on the outer diameter and the tip speed of an impeller is considered, the tip speed decides the pressure ratio of the compressor. Because the ratio of relative velocity and the tip speed at the inlet and the outlet have a certain relationship, the relative velocity ratio at the inlet to the outlet has an optimum value. In addition, the tip speed could not change irrespective of the outer diameter of the impeller since the total pressure ratio was set to 3 in the present design. When the diameter of the impeller is one tenth, Reynolds number also becomes one tenth. Therefore, downsizing of an impeller outer diameter reduces the Reynolds number. On the other hand, from
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Journal of Thermal Science, Vol.12, No. 1, 2003
the point of view of the surface roughness of an impeller, the minimum roughness is decided by the method of manufacturing. Thus, the friction loss due to the increase of relative roughness is very considerable as the decrease of an impeller diameter. For example, in the Moody diagram, when the roughness increases, both the wall friction and the disk friction losses increase. Furthermore, a 2-dimensional impeller will generate large amount of the separation loss due to the steep curvature in the meridian plane at the inlet. The incidence loss also exists even at the design point due to the 2-D configuration, which becomes much larger at the off-design point. In addition, it is difficult to decrease the tip clearance in size for a small impeller. The ratio of the clearance and the impeller outlet width will increase as the decrease of the impeller outer diameter. If the outer diameter of the impeller becomes much smaller, the ultra micro compressor may be operated in laminar flow region. When the Reynolds number based on the relative velocity and impeller chord length is considered, the operating point of the impeller approaches to the laminar flow region even for the ten times model. The operating point sometimes get into the laminar, transient and turbulent flow region. The performance characteristics of an impeller with such kind of operations are not yet clarified including unsteady phenomena such as the surge and the rotating stall. Though the CFD is the useful tool for the design of such impellers, the performance characteristics will not be able to be predicted by the present methodology. Therefore, the study on an ultra micro compressor must be started from establishment of the design methodology. In addition, CFD remains many problems to be solved especially in the transition and the unsteady flow region.
Design of compresssor The designed impeller and the diffuser are shown in Figs.l(a) and l(b). In the design, the inducer incidence, the wall friction, the secondary flow, the leakage flow, the mixing and the disk friction losses were taken into account. The relationship between the blockage at the diffuser throat and the pressure recovery coefficient were also estimated. The Wiesner's formula was used for the slip coefficientI61. The modifications for the main dimensions of the impeller by considering the inlet and outlet total pressure, the static pressure, the relative velocity, the absolute velocity, the flow angle and the temperature were tried. In order to optimize the relative velocity ratio in the impeller, the meridional configuration was changed instead to change the blade thickness in the parallel wall shapes of the hub and the shroud. As described before, it is easily estimated that the attainable highest efficiency of the tested impeller will become much lower than that of conventional
3-dimensional impeller. The operating range of that will be much narrower. Thus, the benefit of 2-dimensional impeller will be only in its productivity.
(a) Impeller i I
/
(b) Diffuser Fig.1 Impeller and diffuser
Prediction of Internal Flow and Performance The numerical analysis was carried out for the impeller. A strong separation was found around the inlet of the impeller as estimated. Fig.2 shows the velocity vector on the mid span of the impeller in the meridian plane. For the simplicity of the calculation, the shape of the leading edge was set straight. The vortices due to the large separated region at the tip and the shroud can be seen. The contour maps of the Mach number on the blade to blade surfaces are shown in Fig.3. The wake due to the separation is concentrated to the shroud side and the velocity was decreased. The separation appeared at the suction surface on the hub side. In the mid span, it was also seen both on the pressure and the suction sides. The
Shimpei Mizuki et al.
Design and Prototyping of Micro Centrifugal Compressor
Fig.2 CFD result of internal flow in mid span
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low value of the efficiency could be estimated by such complicated flows. The inverse design method was also tried to optimize the blade shape of the impeller by using TURBO design F]. One of the results is shown in Fig.4. It is clear that 3-dimensional shape will be better for the shape of the blades. Fig.5 shows an example of the performance characteristics calculated by the prediction method for the compressor performance, where ir and G indicate the total pressure ratio and the mass flow rate, respectively. The equi-efficiency curves are superimposed. The designed mass flow rate and the pressure ratio at n = 220,000 r/rain were G = 0.033 kgf/s and lr = 3. Then, the relative velocity ratio was at 0.68. According to the prediction method used in this study [6], big differences in the results occurred by changing the value of assumed constant such as the wall friction coefficient. The values for the design of 3-dimensional compressor with bigger diameter than that of the present study was used. It will be difficult to expect the accurate prediction of the performance characteristics for the ultra micro compressor with 2-dimensional shape. The main dimensions of the impeller and diffuser are shown in Table. 1.
(a) Shroud side
.d
i,
(b) Mid span Fig.4 Example of compressor by inverse design
50
4.0
30
2.0
1.0~
(c) Hub side
0.01
I
2.19
O0 001
I
0015
i
002
I
I
0.025
0.03
I
0.035
i
I
004 0045 0.05
G (kgf/s)
Mach Number Fig.3 Contour map of Mach number in impeller channel
Fig.5 Prediction of performance characteristics
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Journal of Thermal Science, Vol.12, No.l, 2003
Table 1 Main dimensions of designed compressor
Impeller
Diffuser
Inlet diameter (mm)
20
Outlet diameter (ram)
40
Number of blades
16
Blade thickness (nun)
0.5
Inlet blade height (mm)
6.05
Outlet blade height (mm)
2.4
Inlet blade angle (deg)
50
Outlet blade angle (deg)
30
Outlet flow angle (deg)
76
Inlet diameter (nun)
42
Outlet diameter (mm)
60
Number of blades
16
Blade thickness (ram)
0.5
Blade height (nun)
2.4
(a) Impeller
Prototyping of Compressor From Figs.6(a) to 6(c), the pictures of the prototype of the present ultra micro compressor are shown. The compressor was driven by using a turbocharger for a small automobile engine, which was replaced and rebuilt only the compressor section consisting of the impeller, the diffuser and the casing with the suction nozzle. The machining of the flow channel was carded out by using a 3-axis milling machine. The impeller was made of aluminum alloy (A7075). Though the thickness of the blades was constant along the span in the design stage, a tapered shape with the root thickness 1 mm was given considering the strength at the operating condition. The maximum stress appeared at the leading edge of the hub side by the stress analysis. Therefore, a certain angle was set from shroud side to hub side, which also prevents the steep change of the flow angle. The diffusers with and without vane were made of brass and attached at the flow section from the shroud side. The clearance between the shroud and the casing was 0.3 mm. The flow through the impeller, diffuser and the scroll casing with the constant cross section was lead to the outlet duct and the flow rate was measured by the orifice in the delivery duct. The rotating speed was measured by a photo-electric revolution counter. The performance characteristics were calculated by measuring the inlet and outlet static pressure at the impeller, the static and the total pressure at the diffuser outlet and the compressor outlet air temperature. The cross sectional view of the compressor was shown in Fig. 7.
(b) Diffuser (Vaned)
(c) Casing
Fig.6 Elements of compressor
55
,I
Ng.7 Cross sectional view of compressor assay
Shimpei Mizuki et al.
Design and Prototyping of Micro Centrifugal Compressor
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Experimental Results and Discussion
Conclusion
The performance test was carried out by using cold air. The rotating speed were maintained at n -- 50,000 r/min and n = 70,000 r/min. In the cold air test, the attainable maximum rotating speed was at n = 100,000 r/min. Fig.8 shows the experimental results for the tested compressor with the vaned and the vaneless diffusers. The time-dependent pressure fluctuated when the flow rate slightly reduced from the maximum efficiency, which indicated the occurrence of the surge. From both results of n = 50,000 r/min and n = 70,000 r/min, the vaned diffuser had narrower operating range and gained higher total pressure ratio than the vaneless diffuser. In addition, the surge occurred just after the maximum pressure ratio.
In order to establish the design methodology of an ultra-micro compressor, a ten times size model was designed and tested. Only the preliminary results were reported in this first report. As the combustion test rig will be prepared in the near future, the experiment aimed higher efficiency and wider operation range for a 3-dimensional impeller will be performed.
1,16 1,14
~
~
-'*-" Vaneess(n=50000r/mm) ~ Vaned (n=50000r/rnm)
. "".
.,i..
Vane4ess(n=7OOOOr/mm)
1.12 1.10 k~ 1.08 1.06 1.04 1.02 1,00
I
0,005
I
0.01
I
0,015
G (kgf/s) Fig.8 Performance characteristics (Experimental result)
0.02
Acknowledgement The present study was partially supported by the New Energy and Industrial Technology Development Organization, JAPAN. The authors also appreciate for the Ishikawajima Heavy Industries Co., Ltd. for the support of the strength and vibration analysis of the present impeller.
References [1] Ashley, S. Turbine on a Dime. Mechanical Engineering, 1997 [2] Special Issue of Micro Gas Turbine. Journal of GTSJ, 2001, 29(3) [3] Tanaka, S, et al. Design and Fabrication Challenges for Micromachined Gas Turbine Generators. In: Proc. of the 9th ISROMAC. 2002 [4] Isomura, K, et al. Design Study of a Macromachined Gas Turbine with 3-Dimensional Impeller. In: Proc. of the 9th ISROMAC. 2002 [5] Anzai, A, et al. Design System for High Performance Centrifugal Compressor. Technical Report of EBARA Co., Ltd., 1990, 146:27--35 [6] Galvas, M R. Analytical Correlation of Centrifugal Compressor Design Geometry for Maximum Efficiency with Specific Speed. NASA TN D6729, 1972 [7] TURBO-design -1. Advanced Design Technology Ltd