Journal of Thermal Science Vol.20, No.4 (2011)
312317
DOI: 10.1007/s11630-011-0475-4
Article ID: 1003-2169(2011)04-0312-06
Prototyping of Ultra Micro Centrifugal Compressor-Influence of Meridional Configuration Toshiyuki Hirano1, Tadataka Muto2 and Hoshio Tsujita3 1. Production Systems Engineering Course, Tokyo Metropolitan College of Industrial Technology 2. Graduate School of Engineering, Hosei University, 3-7-2, Kajinocho, Koganei-shi, Tokyo, 184-8584, Japan 3. Department of Mechanical Engineering, Faculty of Science and Engineering, Hosei University © Science Press and Institute of Engineering Thermophysics, CAS and Springer-Verlag Berlin Heidelberg 2010
In order to investigate the design method for a micro centrifugal compressor, which is the most important component of an ultra micro gas turbine, two types of centrifugal impeller with 2-dimensional blade were designed, manufactured and tested. These impellers have different shapes of hub on the meridional plane with each other. Moreover, these types of impeller were made for the 5 times and the 6 times size of the final target centrifugal impeller with the outer diameter of 4mm in order to assess the similitude for the impellers. The comparison among the performance characteristics of the impellers revealed the influence of the meridional configuration on the performance and the similitude of the compressors.
Keywords: Ultra Micro Centrifugal Compressor, Performance Characteristics, Impeller, Meridional Configuration
Introduction Recently several studies for an ultra micro gas turbine have been actively tried to use for portable and reusable electric power sources, ultra micro jet engines and so on [1-3]. There are many challenges toward the practical use of an ultra micro gas turbine and the most significant challenge is to improve the efficiency of ultra micro compressor in order to generate the power output and achieve the gas turbine cycle. However, the design methodology for ultra micro gas turbines has not been established yet. The designed attainable total pressure ratio of these impellers was 3, which was estimated from the cycle study of the present ultra micro gas turbine system. In order to attain the pressure ratio of 3, the required peripheral velocity of impellers was estimated to be about 450-500 m/s. On the other hand, the impeller for an ultra-micro centrifugal compressor is desired to be
designed using 2-dimensional blades taking account of the workability and productivity [4]. However, because the rotational speed becomes extremely high by the decrease of the impeller outlet diameter to attain the pressure ratio of 3, the centrifugal force acting on the root of blades exceeds the material strength, and consequently the operation at the required rotational speed must be impossible. Therefore, in order to apply the 2-dimensional impeller to an ultra micro gas turbine, it is important to optimize the meridional configuration of the impeller by considering the stress acting especially on the root of blades. In this study, in order to examine the effect of the modification of the meridional configuration on the performance of impeller with 2-dimensional blade, two types of impellers with different meridional configuration were designed, manufactured and tested. Moreover, 5 times and 6 times models for these two types of impellers
Received: October 2010 Hirano Toshiyuki: Associate professor www.springerlink.com
Toshiyuki Hirano et al.
Nomenclature b2 D1 D2 G N Pa Ps Psr Pt Q
Prototyping of Ultra Micro Centrifugal Compressor-Influence of Meridional Configuration
impeller outlet blade height [m] impeller inlet diameter [m] impeller outlet diameter [m] mass flow rate [g/s] number of blades [-] atmospheric pressure [Pa] static pressure [Pa] pressure recovery ratio [-] total pressure [Pa] flow rate [m3/s]
t u2 α β ρ Subscript 1 2 3 5
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thickness of blade [mm] impeller outlet peripheral velocity [m/s] impeller inlet blade angle [deg.] impeller outlet blade angle [deg.] density [kg/m3] impeller inlet impeller outlet diffuser outlet compressor outlet
with the outlet diameter of D2= 20 mm and 24 mm for the final target compressor were manufactured and tested. The similitude for impeller was examined by comparing the performance characteristics of 5 times model with those of 6 times model.
Design and Prototyping Impeller In this study, a turbocharger for an automobile engine was employed, in which the centrifugal impeller for the compressor was replaced by the designed impellers which have the 2-dimensional configurations with the outlet diameter D2 of 20 mm and 24 mm. Four types of impellers with different meridional configuration and diameter were designed and manufactured. The main specifications of the impellers are shown in Fig.1 and Table 1. The configuration of shroud casing on the meridional plane in Type-A was designed to satisfy the optimum relative velocity ratio of 0.68 with the hub endwall normal to the rotating axis (Fig.1(a)). In Type-B impeller (Fig.1(b)), the configuration of shroud casing is same as that of Type-A, but that of hub endwall was designed to endure the stress caused by the centrifugal force due to the impeller rotation. Compressor Figure 2 and Table 2 show the cross-section of compressor and the main specification of the compressor, respectively. The configuration of 5 times model is completely analogous to that of 6 times model up to the diffuser outlet. The type of diffuser used in this study is vaneless diffuser and has the radius ratio of 1.8. The test rig is shown in Fig.3.
Experimental Apparatus and Method The performance tests were carried out according to the condition of JIS B 8340. In this study, the cold air
(a) Type-A
(b) Type-B Unit[mm]
Fig.1
Impeller
supplied from a screw compressor was used to drive the turbine rotating with the coaxial impeller. The flow rate was measured by the flow meter. The total pressure at 54 mm downstream of the compressor outlet was measured by a micro pressure sensor installed in the delivery duct. During the tests, the static pressures at impeller outlet, the static pressure and the temperature at diffuser outlet, the static and the total pressures and the temperature at
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Table 1
Dimensions of impellers 6t-model
5t-model
D1
10.8
9.0
D2
24.0
20.0
N
12
T
0.3
b2
1.44
1.20
α
50
β
30
Unit[mm] P1 , P2
Impeller outlet static pressure
P3
Diffuser outlet static pressure
P4
Static pressure in scroll
P5
Compressor outlet static pressure
P6
Compressor outlet total pressure
T3
Diffuser outlet temperature
T5
Compressor outlet temperature
Fig.4
Measurement points
Results and Discussion
Unit[mm]
Fig.2 Table 2
Compressor
Dimensions of Compressor
Symbol in Fig.2
6t-model
5t-model
A[mm]
61.6
54.4
B[mm]
43.2
40.6
C[mm]
40.6
34.0
D[mm]
25.0
21.0
E[mm]
10.8
F[mm]
9.0 13.11
Fig.3
Test rig
compressor outlet were measured as shown in Fig.4. The tested rotational speeds were from 100,000 to 160,000 r/min which were measured by a photo-electric revolution counter.
Influence of meridional configuration The static pressures at the impeller outlet Ps1 and at the diffuser outlet Ps3 and the total pressure at the compressor outlet Pt5 for 5t-model are shown in Fig.5, which are normalized by the atmospheric pressure Pa. In these figures, G is the corrected mass flow rate. Those for 6t-model are shown in Fig.6. For 5t-model (Fig.5), the increase of the rotational speed enhances the pressure ratios for both Type-A and Type-B. The maximum flow rates for Type-A and Type-B are almost the same at each rotational speed. The pressure ratios at the impeller outlet (Fig.5(a)) for Type-A are somewhat higher than those for Type-B at the high flow rate. On the other hand, at low flow rate, the latter are higher than the former. These tendencies are sustained at the diffuser outlet (Fig.5(b)) and also at the compressor outlet (Fig.5(c)). For 6t-model (Fig.6), the pressure ratios for Type-A and Type-B are increased by the increase of the rotational speed as the same as in the case of 5t-model. However, the maximum flow rate for Type-A is higher than that for Type-B unlike 5t-model. The pressure ratio at impeller outlet (Fig.6(a)) is higher for Type-A at high flow rate but for Type-B at low flow rate as also observed in the case of 5t-model. However, at the diffuser outlet (Fig.6(b)) and the compressor outlet (Fig.6(c)), the pressure ratios for Type-A are higher than those for Type-B at all flow rates. As mentioned above, the pressure ratio is higher for Type-A at high flow rate but for Type-B at low flow rate
Toshiyuki Hirano et al.
Prototyping of Ultra Micro Centrifugal Compressor-Influence of Meridional Configuration
(a) Static pressure ratio at impeller outlet
(a) Static pressure ratio at impeller outlet
(b) Static pressure ratio at diffuser outlet
(b) Static pressure ratio at diffuser outlet
(c) Total pressure ratio at compressor outlet
(c) Total pressure ratio at compressor outlet
Fig.5
Performance characteristics of 5t-model
from the impeller outlet up to the compressor outlet for 5t-model (Fig.5) but only at the impeller outlet for 6t-model (Fig.6). These results would suggest that there is a difference in the diffuser performance between 5t-model and 6t-model. So, in order to assess the diffuser performance for each impeller, the pressure recovery ratios of diffuser defined by Eq.(1) were evaluated and shown in Fig.7. P P Psr s 3 2 s1 (1) u2 2 In the case of 5t-model (Fig.7(a)), there is not remarkable difference between the pressure recovery ratios for Type-A and Type-B. However, in the case of 6t-model (Fig.7(b)), the pressure recovery ratio for Type-A is higher than that of Type-B and exhibits the highest performance among all cases. These indicate that the merid-
Fig.6
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Performance characteristics of 6t-model
ional configuration of impeller does not influence the performance characteristics of the impeller itself but affects the combination of the impeller with the diffuser. Moreover, it is thought that the highest performance of the diffuser results in the enhancement of the maximum flow rate for Type-A in 6t-model. Similitude of compressor Figure 8 shows the comparisons of the performance curves of Ps1, Ps3 and Pt5 for Type-A at the rotational speed of 120,000r/min in the case of 5t-model with those at 100,000r/min in 6t-model, which have the same impeller outlet peripheral velocity. The flow coefficient φ in Fig.8 is defined as follows. Q (2) D2b2u 2 Those for Type-B are given in Fig.9.
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(b) 6t-model
(a) 5t-model
Fig.7
Pressure recovery ratio at diffuser
(a) Static pressure ratio at impeller outlet
(a) Static pressure ratio at impeller outlet
(b) Static pressure ratio at diffuser outlet
(b) Static pressure ratio at diffuser outlet
(c) Total pressure ratio at compressor outlet
(c) Total pressure ratio at compressor outlet
Fig.8 Performance characteristics of Type A
Fig.9
Performance characteristics of Type B
Toshiyuki Hirano et al.
Prototyping of Ultra Micro Centrifugal Compressor-Influence of Meridional Configuration
The maximum flow coefficient in the case of 5t-model is higher than that in 6t-model for both of Type-A (Fig.8) and Type-B (Fig.9). This would be caused by the use of the same size of chambers and delivery ducts both in 5t-model and 6t-model. On the other hand, the pressure ratios from the impeller outlet up to the compressor outlet in 6t-model possess higher value than those in 5t-model for both Type-A and Type-B. This would be caused by the tip-clearance size which is relatively larger in 5t-model than in 6t-model. The pressure ratio at the impeller outlet in 5t-model is almost the same as in 6t-model for both Type-A (Fig.8(a)) and Type-B (Fig.9(a)). Therefore, the similitude of performance characteristic is thought to be maintained up to the impeller outlet because of the slight difference in size between 5t-model and 6t-model. In the comparison of the pressure ratios at the diffuser outlet (Fig.8(b), 9(b)) and the compressor outlet (Fig.8(c), 9(c)) in 5t-model with those in 6t-model, the difference obviously appears at the diffuser outlet and the compressor outlet for Type-A. This is caused by the difference in the diffuser performance mentioned before.
Conclusions The following conclusions were derived from the present study. (1) The pressure ratio at the impeller outlet is higher for Type-A at high flow rate and higher for Type-B at low
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flow rate. (2) The meridional configuration of impeller does not influence the performance characteristics of the impeller itself but affects the combination of the impeller with the diffuser. (3) The similitude of performance characteristic is maintained up to the impeller outlet because of the slight difference in size between 5t-model and 6t-model.
References [1] Epstein, A. H., et al.：Micro-Heat Engine and Rocket Engines, -The MIT Microengine Project-, AIAA Paper, 97-1773, Snowmass Village, Colorado, 1997. [2] Epstein, A. H.: Millimeter-Scale, MEMS Gas Turbine Engines, ASME Turbo Expo 2003, GT-2003-3886, Atlanta, 2003. [3] Isomura, K., Teramoto, S., Hikichi, K., Endo, Y., Togo, S. and Tanaka, S., The performance evaluation of 10 mm micro-centrifugal compressor, 11th Power & Energy Technology Symposium Report No.06-8, JSME, Tokyo, 2006. [4] Hirano, T., Yamaguchi, N., Minorikawa, G., Tsujita, H. and Mizuki, S., Design and Prototyping of Micro Centrifugal Compressor for Ultra Micro Gas Turbine, 31th Gas Turbine Society of Japan Report, pp.177-182, Kitami, 2003.
