Journal of Thermal Science Vol.24, No.4 (2015) 313322
DOI: 10.1007/s11630-015-0790-2
Article ID: 1003-2169(2015)04-0313-10
Non-axisymmetric Flow Characteristics in Centrifugal Compressor WANG Leilei1,2, LAO Dazhong1*, LIU Yixiong1, YANG Ce1 1. School of Mechanical Engineering, Beijing Institute of Technology, Beijing 100081, China 2. College of Mechanical and Electrical Engineering, Hebei University of Engineering, Handan 056038, China © Science Press and Institute of Engineering Thermophysics, CAS and Springer-Verlag Berlin Heidelberg 2015
The flow field distribution in centrifugal compressor is significantly affected by the non-axisymmetric geometry structure of the volute. The experimental and numerical simulation methods were adopted in this work to study the compressor flow field distribution with different flow conditions. The results show that the pressure distribution in volute is characterized by the circumferential non-uniform phenomenon and the pressure fluctuation on the high static pressure zone propagates reversely to upstream, which results in the non-axisymmetric flow inside the compressor. The non-uniform level of pressure distribution in large flow condition is higher than that in small flow condition, its effect on the upstream flow field is also stronger. Additionally, the non-uniform circumferential pressure distribution in volute brings the non-axisymmetric flow at impeller outlet. In different flow conditions, the circumferential variation of the absolute flow angle at impeller outlet is also different. Meanwhile, the non-axisymmetric flow characteristics in internal impeller can be also reflected by the distribution of the mass flow. The high static pressure region of the volute corresponds to the decrease of mass flow in upstream blade channel, while the low static pressure zone of the volute corresponds to the increase of the mass flow. In small flow condition, the mass flow difference in the blade channel is bigger than that in the large flow condition.
Keywords: centrifugal impeller; static pressure distortion; non-axisymmetric flow; mass flow; blade loading
Introduction For the centrifugal compressor equipped to the vehicle turbocharger, the diffuser is connected with the volute, and the tongue exists unavoidably due to the spiral flow channel of the volute. This asymmetric structure of the volute has significant effects on the flow field of the diffuser and impeller[1,2]. It affects the distribution of the pressure and the velocity at the circumferential direction inside the volute. For the vaneless diffuser, these distortions have great influence on the flow field at impeller outlet[3], which propagates reversely against the mainstream to the leading edge of the impeller[4], thus, the non-axisymmetric flow is formed in the centrifugal com-
pressor. Additionally, the aerodynamic performance and the operating stability of the compressor can also be reduced due to the effect of the volute on the impeller[5]. Related studies have found that the secondary flow in the volute plays the major role on the pressure distortion and velocity distribution in the volute, furthermore, the three-dimensional flow field characteristic in the volute should be also considered[6-8]. Gu, et al[9] investigated the interference between the volute and the impeller under off-design condition with numerical method. The results confirmed the presence of static pressure and total pressure distortion at the diffuser outlet and found these two forms of pressure distortion have phase difference at the circumferential direction and propagate upstream along
Received: March 2015 LAO Da-zhong: Associate professor This research was sponsored by the National Natural Science Foundation of China (No.51276017). www.springerlink.com
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vaneless diffuser. Zheng, et al[10] conducted the investigation on a high-pressure ratio transonic centrifugal compressor in off-design operating condition with numerical method, it pointed out that the volute severely affects the flow stability of the compressor and limits the flow stability range below the design flow condition. Subsequently, Zheng conducted further experimental research on the compressor. The static pressure variations at the volute, impeller outlet, the leading edge of the main blade and the leading edge of splitter blade were measured. The results showed that the circumferential asymmetry of the volute causes the obvious asymmetric flow in the compressor. The static pressure distortion of the diffuser occurs at about 90° circumference position of the volute tongue, and the distortion degree is slightly influenced by the rotational speed of the impeller [11]. However, in their work, the "frozen rotor" was adopted in the numerical model, and it could not predict the decay of the pressure distortion when it propagates. Actually, due to the rotation of the impeller, the effect of the pressure distortion at the impeller outlet on the upstream flow is often smaller than that by the steady calculation [9]. Previous studies just pointed out the location of the pressure distortion in the volute and the pressure distribution at the impeller inlet along the circumferential direction, but they did not focused on the unsteady characteristics of the circumferential distribution of mass flow among blade channels. Meanwhile, a few works involved the reverse propagating phenomenon of the non-uniform pressure distribution from the volute to the impeller as well as the relevant research on the blade loading changes caused by this reverse propagating phenomenon. the non-axisymmetric flow characteristics in centrifugal compressor and the reverse propagating characteristic of the static pressure distortion in volute were investigated in this work by experimental and numerical method. Firstly, the circumferential static pressures at shroud wall and the compressor inlet were measured. Then, the static pressure distribution on the shroud wall at typical circumferential position of the volute was analyzed by combining with the unsteady numerical results. Meanwhile, the non-axisymmetric flow characteristics of the compressor caused by the circumferential static pressure distortion were investigated in this work. It mainly focused on the reversed propagating process of the static pressure distortion in volute and its effects on the mass flow in blade channels, impeller inlet conditions and the blade loadings.
Numerical Model A centrifugal compressor for vehicle turbocharger has been chosen as the research object in this paper. The geometric parameters are shown in Table 1.The design
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rotational speed of this compressor is 80 kr/min. The numerical method is referred by Ref. [12]. The spatially discrete grid of the cascade passage is generated by using the IGG/Autogrid software package. The flow field domain of the cascade passage is meshed by the O4H type structured grid. H-type block is adopted to mesh the extension section and the diffuser. O-type block is used to surround the blade row. The grids of the inlet pipe and the volute are generated manually using the IGG software package and the grid quality are enhanced by butterfly grid technology. The y+ value in most region of the computation domain is less than 5. The Multi-grid method is also adopted to speed up the convergence of the numerical solution. The commercial software named FINE/TURBO is used to solve the three-dimensional Reynolds-averaged Navier-Stokes equation. In this paper, the SpalartAllmaras model is adopted. The flow solver is based on a cell centered finite volume approach, with the space discretization by central difference scheme, the time integration by the four order Runge-Kutta explicit method, and the numerical step is discretized by the dual time stepping. The grid model adopts the full circle of the impeller. In steady calculation, the "rotor freezing" method is used to treat with the interface between the rotor and stator. The inlet conditions are set by the uniform total temperature, total pressure and the airflow direction is axial. In large mass flow condition, the outlet of the compressor is set by the mean static pressure. When the operating condition of the compressor approaches to the surge point, the outlet is set by the mass flow. For the unsteady computation, the interface between the rotor and stator is treated with the "sliding mesh" method. The physical time step is 175 within a rotor rotational period, which is corresponding to the time step of about 4.286μs. Table 1
Main geometry parameters of the compressor Item
Unit
Value
Main blade number
−
7
Splitter blade
−
7
Impeller exit backswept angle
°
35
Design speed
Kr/min
80
Impeller inlet diameter D1
mm
61
Impeller exit diameter D2
mm
90
Diffuser inlet diameter D3
mm
108
Volute inlet diameter D4
mm
144
Experiment The experimental work was carried out at compressor test rig at turbomachinery institute in Beijing Institute of Technology. The compressor was driven by a turbine. The turbine was driven by the airflow from a compressed
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air supply source. By adjusting the valve on the experimental pipeline, the different flow condition points were got with the constant rotational speed. In the experiment process, the performance parameters, such as the mass flow, rotational speed, temperature and pressure at compressor inlet and outlet were measured. For the convenience of analysis, the main impeller blades of the compressor were numbered and the circumferential reference angles of the volute were defined as shown in Fig.1, the position of volute tongue was corresponding to the circumferential angle of about 50°. Besides the measurement of performance parameters of the compressor, the pressure on the impeller inlet and the pressure distribution from the impeller inlet to the diffuser were also measured. the positions of the measure points were shown in Fig.2. For the pressure distribution of the impeller inlet, 6 measure points (C1~C6) were arranged along the circumferential position at the impeller upstream with the distance of 25 mm to the leading edge of the main blade. The circumferential position of the measuring point C1 corresponded with the reference position of the circumferential 0° of the volute. For the pressure distribution on the shroud wall, 12 measuring points (P1~P12) were arranged at the circumference position of the volute, and the circumference reference angle was 90°. Among them, P1 was set at the 5 mm upstream in axial direction from the leading edge of the main blade; P2 and P4 were set at the leading edge of the main blades and splitter blades respectively; P12 was arranged inside the diffuser 1.4R2.
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Results and Analysis Test Results and Numerical Validation Fig.3 shows the comparison of the experimental and the calculated performance curve of the compressor, wherein, Fig.3(a) is the pressure ratio and flow, Fig.3(b) is the efficiency and flow characteristic.
Fig. 3 Aerodynamic performance comparisons between the test and calculation value
Fig. 1
Main blade number and the circumferential position
Fig. 2 Locations of the measure points
As shown in Fig.3(a), for the rotating speeds of 60 kr/min and 70 kr/min conditions, the pressure ratio and flow curve are in well uniformity for the test and numerical calculation value. For the 80 kr/min condition, some differences appear between the measurement and calculation value in large mass flow condition, the maximum error is about 4%. In small mass flow condition, the pressure ratio and the flow curves match well. For the surge flow condition, there are some differences between the calculation results and the experimental values, which mean that using the steady calculation can not capture the surge point of the compressor accurately. It can be seen from Fig.3(b), the calculation efficiency of the compressor are lower than that of the experimental value, but they gradually conformed to each other with the flow rate increasing. The trends of the efficiency in the whole mass flow range of the compressor are in well conformity, and
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the error is less than 2.85% within acceptable range. Static Pressure Distribution in the Volute The further analysis in this paper is based on the design speed of the compressor. Relevant studies have showed that the airflow near the volute tongue would appear strong non-uniformity in off-design mass flow condition[4]. Therefore, two typical off-design flow conditions are adopted in this paper. The larger mass flow is 0.40 kg/s, which is close to the choke point; and the smaller mass flow is 0.26 kg/s near the surge point, as shown in Fig.3(a). It is essential to analyze the non-axisymmetric flow phenomenon in the volute because the non-symmetrical structure of the volute is the fundamental cause of the non-axisymmetric flow in compressor. The non-symmetrical volute induces its inner flow with the non-axisymmetric characteristic, thus makes the circumferential static pressure distortion induced in volute. Generally, when the compressor is close to the surge point, the static pressure along the flow direction in the volute increases. When the compressor is close to the choke point, the static pressure along the flow direction in the volute decreases. References [1,3,9,11,14] investigated the static pressure circumferential distribution of the diffuser and the volute of the centrifugal compressor by either experimental test or steady numerical calculation method with the conclusion that the variation of the static pressure is consistent with the above discussion. Fig.4 presents the time-averaged static pressure distribution on the intermediate longitudinal section of the volute with two different operating conditions. It can be found that the circumferential static pressure distortion exists inside the volute with a high static pressure area near the tongue. The size of the high static pressure area changed with the variation of the mass flow. This can be explained by the blocking effect of the tongue against the airflow passing near the volute tongue, which means that the volute tongue is one of the most important factor for the non-axisymmetric flow inside the compressor. It can also be seen that the static pressure decreases continuously from the tongue to volute outlet along the flow direction in large flow condition, which indicates that the flow within this region is the accelerated flow. For the small flow condition, the static pressure distribution in the volute is just on the contrary of the large flow condition, the static pressure from the tongue to the volute outlet increases and the flow is the decelerated flow. According to the static pressure distribution, it can be found that the blocking effect of the tongue on the flow at the impeller exit with small flow condition is more obvious than that in large flow condition. For the large flow condition, there is low static pressure region in the circumferential locations of 270°~360°
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of the diffuser inlet due to the influence of the geometric change of the volute as shown in Fig.4(a) marked with A. For the small flow condition, the static pressure circumferential distribution of the diffuser inlet also exists non-axisymmetric phenomenon. there are two low pressure regions, one of which locates at the circumferential location of 230°~300°, as shown in Fig.4(b) marked with B, the other is at the circumferential position of 60°~120° marked with C in Fig.4(b). Some related studies also found the low static pressure region inside the diffuser appear at the circumferential position corresponding to C region in the small flow condition, however, the existence of low static pressure at region B has not been focused on [11,14]. For the region B, in the process of the circumferential motion of the airflow from the tongue to the compressor outlet, it can be regarded that the effect of the centrifugal force causes the high static pressure region at the outside of the volute and the low static pressure at the impeller outlet. Meanwhile, with the impeller rotating, the intake mass flow in volute passage increases gradually. Thus, the centrifugal force is furtherly enhanced, which is the principle reason for the formation of the low static pressure region. But behind the circumferential angle of 270°, the flow separation appears at the outlet of the volute, then the centrifugal force decreases and the static pressure in the volute increases again.
Fig. 4 Time-averaged static pressure distribution in volute, 50% diffuser width
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Non-axisymmetric Static Pressure Distributions in Diffuser and Impeller The circumferential non-uniform static distribution of the volute can inevitably cause the non-asymmetry for flow field of the impeller and volute, this non-axisymmetric distribution could be observed for the static pressure difference at the shroud wall of the compressor impeller and the static pressure difference among different circumferential positions in diffuser. The comparisons of the static pressure in the volute for the experimental and calculation data at the circumferential position of 90 are shown in Fig.5. In these figures, the abscissa axis is the dimensionless position of the measurement points; the ordinate axis is the dimensionless static pressure value. The calculation results are given including the mean static pressure and the unsteady calculation results. It can be seen the experimental static pressure is larger than that of the calculated time-averaged value at the middle chord position of the impeller, and the difference is about 7%, which is within the range of the unsteady static pressure fluctuation. In small flow rate condition, the experimental results of the static pressure at each measuring point of the shroud agree well with the calculated values.
Therefore, it is supposed that the numerical calculation can reflect the variation of the static pressure from the impeller inlet to the volute inlet along circumferential position. As is shown in Fig.5 for the unsteady calculation results, the pressure fluctuations at measuring points also can be observed due to the impeller rotating. It is shown that the unsteady fluctuation of the static pressure at the leading edge of the main blade is much larger in large flow condition, marked with the dotted line in Fig.5. In small flow rate condition, the fluctuation of the static pressure at the leading edge of the main blade is relatively small with the lower static pressure fluctuation. Fig.6 shows the static pressure distributions at four circumferential positions of the shroud of the impeller comprised 90°, 180°, 270° and 360° respectively with the two operating conditions. From the impeller inlet, the difference of the circumferential distribution of the static pressure come to appear, and in large flow condition, the difference is larger than that of the small flow condition. For the large flow rate condition, from the splitter blade inlet to diffuser outlet, the maximum of the static pressure is at circumferential 90° position, the minimum value locates at the circumferential 360° position. Combined with Fig.5(a), it can be found that the circumferential pressure difference between the diffuser and impeller is
Fig. 5 Static pressure variation at measured point of the compressor shroud
Fig. 6
The static pressure variation of the shroud at different circumferential positions
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caused by the differences of the circumferential pressure distribution of the volute. As the above discussion, for the large flow rate condition, with the circumferential angle increasing, the static pressure in the volute decreases, which results in the maximum static pressure appears at circumferential 90° position and the minimum locates at circumferential 360°. For the small flow rate condition, in most region of the impeller and diffuser, the static pressure at circumferential 90° position is the minimum value, the maximum value locates at the circumferential 360° position, and the static pressure at circumferential 270° position is less than that of circumferential 180° position, as shown in Fig.5(b). By comparing the differences of the circumferential pressure distribution, it is concluded that the pressure difference of the volute has significant effects on the circumferential pressure distribution in the impeller and diffuser. For large flow rate condition, the circumferential pressure difference of the volute is large, and the circumferential pressure difference in the impeller and diffuser is also large. While in the small flow rate condition, the circumferential static pressure difference of the volute is small, so is the static pressure difference in the impeller and diffuser. It also can be seen that in Fig.6, in the large flow rate condition, the circumferential static pressure difference of the volute has a little effect on the leading edge of the main blade. In the small flow rate condition, this pressure difference of the volute has no effect on the leading edge of the main blade, namely, the static pressure circumferential distribution of the impeller inlet approximately keeps the same under the small flow rate condition. For the leading edge of the splitter blade, in small flow rate condition, the static pressure close to 180° circumferential position is significantly lower than the other circumferential positions. In the large flow rate condition, the static pressure at the leading edge of the splitter blade changes significantly, and the static pressure at 180° circumferential position is the maximum. For the static pressure distribution at leading edge of the blade, when in the small flow rate condition, the surge flow of the compressor could be reduced by non-axisymmetric groove structure at the leading edge of the splitting blade and the impeller inlet, but the form of the non-axisymmetric structure is hard to be determined. Considering the significant difference of the circumferential distribution of the static pressure at the leading edge of the splitter blade between the large flow condition and the small flow rate condition, and the large fluctuation amplitude of the static pressure (see Fig.5(a)), it is supposed that the asymmetric groove structure may significantly reduce the blocking flow of the compressor. To study the effect of the circumferential static pressure on the compressor inlet, it is presented in Fig.7 that the experimental and computational static pressure of the
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measuring points at the compressor inlet in two operating conditions, of which the calculation results include the time-averaged static pressure of the measuring point and its fluctuation with time. It can be seen that the circumferential non-uniform of the static pressure distribution appears at the measure point in large flow condition both from the experiment and numerical calculation results, but the difference is not significant. In small flow condition, the static pressure at compressor inlet along the circumferential direction is basically unchanged. It indicates that the circumferential pressure distribution of the compressor volute has a little effect on the static pressure distribution at the impeller inlet. Especially in the small flow condition, this influence almost can be neglected.
Fig. 7 The circumferential distribution of the static pressure at compressor inlet
The calculation results in Fig.7 also show the unsteady characteristics of the circumferential pressure distribution at impeller inlet. The fluctuation of the static pressure in small flow rate condition is larger than that in the large flow condition. Reverse Propagation of the Circumferential Static Pressure Distortion The circumferential static pressure distortion of the volute affects the static pressure distribution of the impeller outlet. As is shown in Fig.8 the circumferential distribution of the static pressure on the first time step of unsteady calculation at impeller outlet R/R2 =1.045 section, in the two flow rate conditions, TP represents the circumferential position of the volute tongue. The circumferential change trend of the static pressure at impeller outlet is basically the same with the circumferential distribution of the static pressure in volute as shown in Fig.4. But there is the static pressure fluctuation along the circumferential direction. The pressure fluctuation is mainly caused by the blade wake. In large flow condition, the pressure fluctuation is more obvious than that in small flow condition because of the larger velocity loss and the stronger jet-wake interference. Additionally, be-
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Non-axisymmetric Flow Characteristics in Centrifugal Compressor
cause of the existence of mainstream wake, blade wake, jet flow and tip leakage flow at impeller outlet, the static pressure distribution at the impeller outlet along the hub to the shroud is also non-uniform. The change of static pressure at the impeller outlet would certainly cause the variation of the airflow velocity of impeller outlet. It can show the variation of the airflow velocity at the impeller outlet quantitatively through
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the velocity triangle of different circumferential positions at impeller outlet. The outlet airflow in the channel between the main blade 1 and 2 is chosen as the object, and the radial and tangential average velocities at this cross section are made to construct the velocity triangle. The velocity triangles at 90°, 180°, 270° and 360° circumferential position are shown in Fig 9.
Fig. 8 The circumferential distribution of the static pressure at compressor outlet
Fig. 9 Velocity triangle at impeller outlet, R/R2=1.045
It can be seen from Fig.9 that there is obvious difference of the velocity triangle of the impeller outlet at different circumferential positions. As shown in Fig 9(a), for the large flow rate condition, at circumferential 90° position, the radial velocity of the airflow at the impeller outlet is the lowest, and then it gradually increases during the process of impeller rotation from 90° to 360°.The tangential velocity at 90° is relative low, then it increases with the angle increasing, and it reaches the maximum at 180°, then it gradually decreases until 360°. In Fig 9(b), for the small flow rate condition, at circumferential 90° position, the radial velocity of the airflow at the impeller outlet is the highest, and then it gradually decreases, the minimum locates at circumferential 360° position with the angle increasing. There is velocity rise at circumferential 270° position. But the variation of the tangential
velocity is small. For the two conditions discussed above, with the angle increasing from circumferential 90° to 360°, which is corresponding to the airflow propagating from the rear of the volute to the tongue outlet, the absolute flow angle of the impeller outlet (α2) decreases gradually in the large flow rate condition, as is helpful to reduce the transportation distance of the airflow in the diffuser and reduce friction loss. But for small flow rate condition, the absolute flow angle of the impeller outlet increases at the beginning, then decreases, and then increases, which may cause much complex airflow in the diffuser. To illustrate the reverse propagation of the circumferential distribution of the static pressure of impeller outlet along the blade channel to upstream, it is shown in Fig.10 the time-space distribution of the static pressure between
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the pressure surface of the main blade 1 and the suction surface of its adjacent splitter blades along the 90% blade span. In these figures, the abscissa is the dimensionless length of impeller flow field along the meridional direction, the vertical coordinate is dimensionless rotation cycle, ML, SL represents the leading edge of the main blade and the splitter blade respectively, the arrow represents the propagation direction of the high static pressure peak to the upstream, the dotted line represents the 4 typical circumferential position of the impeller. It can be seen from Fig.10(a), for the large flow rate condition, when the impeller outlet is at the circumferential position of the volute tongue, the static pressure of impeller outlet reaches the maximum value, and the propagation strength of the static pressure along the blade channel to upstream is the strongest, it approximately propagates to the impeller inlet at about 180° circumferential position. With the circumferential angle increasing, the static pressure at impeller outlet is gradually decreasing, and its propagate ability along the channel to the leading edge of the blade is also weakening. As shown in Fig.10(b), for small flow rate condition, the highest static pressure of the impeller outlet is at the circumferential position of the volute tongue, while the static pressure at
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the tongue rear zone is the lowest. From this low pressure position, the static pressure at the impeller outlet increases and then decreases with the circumferential angle increasing, and the second low static pressure zone appears at the 270° circumferential position, then it continues to increase along the circumferential direction, which causes two static pressure peaks appear at impeller outlet when it propagates upstream. For the large flow condition, the high static pressure at impeller outlet more concentrated and the circumferential difference of the static pressure distribution is obvious, so that the effect of the high static pressure on the impeller upstream is more prominent. For small flow conditions the circumferential difference of the static pressure is small. The changes of the static pressure and the radial velocity at impeller outlet have effects on the flow capacity in each impeller channel, so the circumferential distribution of the compressor flow is non-uniform. To compare the mass flow change in each impeller channel with two flow rate conditions, the mass flow in the blade channel is represented by the flow ratio, which is defined as the ratio of the mass flow in the impeller channel to the total time-averaged mass flow, written as follows: m m macor 7 (1) where， m represents the mass flow ratio, m is the mass flow in the single impeller channel, macor is the total mass flow. It is shown in Fig.11 the change of flow ratio in single blade channel, the abscissa is the circumferential reference angle of the impeller, the circumferential position of the blade channel is represented by the circumferential angle of the trailing edge of the splitter blade. It can be seen from Fig.11, for the large flow rate condition, when the blade channel rotates to 60° circumferential position, the flow in this channel is the minimum. Along the circumferential direction, the flow of the blade channel increases gradually, and it reaches the maximum at 360° circumferential position. For the small flow rate condition, the flow of the blade channel reaches the maxi-
Fig. 10 The time-space distribution of the static pressure in blade channel, 90% span
Fig. 11 The variation of the flow rate in blade channel
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Non-axisymmetric Flow Characteristics in Centrifugal Compressor
mum at 90°circumferential position, and the second flow peak appears at 240° circumferential position. It is noteworthy that the minimum flow ratio change in blade channel does not appear at the high static pressure region near the front area of volute tongue, but at 150° circumferential position. For large flow rate condition, the difference of the mass flow distribution of each blade channel in different circumferential position reaches 14%, and the difference in small flow rate condition reaches 21%. In the large flow condition, the non-uniform flow distribution in blade channel is stronger than that in small flow rate condition.
Conclusions The circumferential static pressure distribution in the volute of the centrifugal compressor is characterized by non-axisymmetric, especially obvious in the large flow rate condition. The reverse propagation of the static pressure in the volute leads to the non-axisymmetric flow in the compressor. Accordingly, two typical off-design operating conditions (the large and small flow condition) were chosen to investigate the flow characteristics inside the compressor. It can be concluded as follows: The non-axisymmetric static pressure distribution in the volute causes the flow field difference at impeller outlet. For large flow rate condition, the static pressure at different circumferential positions are significantly varying, and the minimum static pressure locates at the 0° circumferential position. In small flow condition, the circumferential pressure difference reduces significantly, but at the 0° circumferential position, the static pressure is the maximum. Therefore, it should be focused on different non-axisymmetric static pressure distributions in large and small flow conditions for the groove structure design at compressor impeller inlet. The reverse propagation path of the pressure fluctuation generated in the volute could be determined with the help of the space-time diagram. Due to the non-uniform degree of the static pressure in the volute, for large flow condition, the strong perturbation pressure wave at the impeller outlet 90° circumferential position can reverse propagate to the impeller inlet about 180° circumferential position; for small flow rate condition, the disturbance pressure wave at impeller outlet is weak, the reverse propagation almost has no influence on the impeller inlet. But for both the two flow conditions, the static pressure distributions in the impeller change along the full circumferential range, so the pressure distributions in the compressor are circumferential asymmetrical. The non-uniform circumferential pressure distribution in volute affects the flow parameters at impeller outlet. For the impeller outlet, in the large flow rate condition,
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the maximum absolute flow angle locates at the circumferential position corresponding volute tongue, along the rotation direction of the impeller, the absolute flow angle decreases gradually; in the small flow rate condition, the absolute flow angle at volute tongue circumferential position is the minimum, the absolute flow angle decreases firstly, then increases, and then decreases, along the impeller rotation direction. This marks the flow direction of impeller outlet with non-axisymmetric characteristics in full circumferential range. The non-axisymmetric flow characteristic in the impeller is also reflected in the mass flow distribution. The high static pressure region of the volute corresponds to the decrease of mass flow decreasing in upstream blade channel, and the low static pressure region of the volute to the mass flow increasing in upstream. In large flow rate condition, the difference of the mass flow distribution of each blade channel in different circumferential position reaches 14%, and the difference in small flow rate condition reaches 21%.
Acknowledgements This research was sponsored by the National Natural Science Foundation of China (No.51276017).
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