The change of the inlet geometry of a centrifugal compressor stage and its influence on the compressor performance

Journal of Thermal Science Vol.22, No.3 (2013) 197−208

Received: January 2013 YANG Ce: Professor This research was sponsored by the National Natural Science Foundation of China (No.51276017) and Ph.D. Programs Foundation ofMinistry of Education of China (No. 20111101130002).

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DOI: 10.1007/s11630-013-0613-2 Article ID: 1003-2169(2013)03-0197-12

The Change of the Inlet Geometry of a Centrifugal Compressor Stage and its Influence on the Compressor Performance

WANG Leilei, YANG Ce, ZHAO Ben, LAO Dazhong, MA Chaochen, LI Du

School of Mechanical Engineering, Beijing Institute of Technology, Beijing 100081, China

© Science Press and Institute of Engineering Thermophysics, CAS and Springer-Verlag Berlin Heidelberg 2013

The impact on the compressor performance is important for designing the inlet pipe of the centrifugal compressor of a vehicle turbocharger with different inlet pipes. First, an experiment was performed to determine the com-pressor performance from three cases: a straight inlet pipe, a long bent inlet pipe and a short bent inlet pipe. Next, dynamic sensors were installed in key positions to collect the sign of the unsteady pressure of the centrifugal compressor. Combined with the results of numerical simulations, the total pressure distortion in the pipes, the pressure distributions on the blades and the pressure variability in the diffuser are studied in detail. The results can be summarized as follows: a bent pipe results in an inlet distortion to the compressor, which leads to per-formance degradation, and the effect is more apparent as the mass flow rate increases. The distortion induced by the bent inlet is not only influenced by the distance between the outlet of the bent section and the leading edge of the impeller but also by the impeller rotation. The flow fields in the centrifugal impeller and the diffuser are in-fluenced by a coupling effect produced by the upstream inlet distortion and the downstream blocking effect from the volute tongue. If the inlet geometry is changed, the distributions and the fluctuation intensities of the static pressure on the main blade surface of the centrifugal impeller and in the diffuser are changed accordingly.

Keywords: centrifugal compressor; inlet pipe; distortion; main blade; diffuser

Introduction

Centrifugal compressors are widely used in military and civil fields and are the most common air compres-sion systems that are applied in turboblower and micro-miniature turbojet/turbofan engines. The flow field of the compressor inlet is usually assumed or assured to be uniform and stable in the design or the test process, re-spectively. In fact, distorted and non-uniform flows at the impeller inlet exist in many application circumstances. For strict limitations of the volume, weight and frame-works in the machines，a bent pipe is often adopted in the inlet of the centrifugal compressor, which may result in

an inlet distortion that affects the aerodynamic perform-ance of the compressor and leads to blade reliability deg-radation.

Therefore, a detailed study of the influence of geomet-ric changes in the inlet pipe on a centrifugal compressor can provide a reference for the design process of the inlet pipe. Some scholars have conducted correlation studies with important conclusions. An unsteady calculation of the inlet distortion in front of the compressor impeller shows that the blade loading has the maximum fluctua-tion range in the leading edge. This range decreases as the flow decreases[1,2]. Engeda[3] used experiments and numerical simulation calculations to study the perform-

198 J. Therm. Sci., Vol.22, No.3, 2013

ance of a compressor when the inlet varies with different pipes, including a straight pipe and bent pipes. The re-sults indicated that a compressor with a 90-deg curved pipe can cause inlet flow distortion. If the inlet configu-ration is changed by adding vanes in appropriate posi-tions in the bent pipe, the performance of the compressor and diffuser can be improved by lowering the secondary flow and providing as uniform a flow as possible for the compressor.

In recent decades, researchers in China have studied the inlet distortion of a centrifugal compressor. Zhou [4] performed experiments to explore the manner in which the inlet radial distortion affects the centrifugal com-pressor performance. He concluded that the comprehen-sive effect is favorable to the operability of the compres-sor if a radial distortion network is installed in front of the inlet guide vane. The tested pressure ratio and effi-ciency are higher than those parameters with no distor-tion, and the surge margin is larger in this case. The pos-sibility that the compressor has a distortion tolerance capacity was also considered. Zhang [5] connected the centrifugal compressor inlet with straight pipes of differ-ing lengths and diameters. He then studied the impact on the compressor performance of these pipes using steady numerical simulations. The whole stage calculation in the compressor including the outlet pipe, a vaneless diffuser and a volute was performed and indicated that the uncor-related performance curve varies dramatically with dif-ferent inlet pipes.

The previous results[6-7] obtained by our research team showed that the inlet flow distortion from a bent pipe diminishes the compressor performance based on the different position at which the bent pipe is mounted. In addition, through studying the unsteady characteristic, it was discovered that the inlet distortion leads to a higher pressure fluctuation before the impeller inlet. The spec-trum structure in the upstream impeller flow field was also changed, which also significantly improved the dis-turbance intensity induced by the rotor’s fundamental frequency. Meanwhile, the blade vibration and tongue region noise are strengthened due to the distortion, the de-gradation of which is more obvious at a low mass flow rate.

In previous studies, researchers have seldom paid at-tention to the alteration of the flow field in the centrifugal compressor inlet, the static pressure distribution on the blade surface and the pressure fluctuation inside the dif-fuser influenced by the flow distortion with different bent inlet pipes. Therefore, this paper presents an investigation of these problems. We also discussed the change of the flow distortion created by the inlet pipe geometry.

Research object

The research object is the centrifugal compressor that is used for a J90S turbocharger, the geometric parameters

of which are shown in Table 1. The original compressor includes a centrifugal impeller, a paralleled-wall vaneless diffuser and a volute. We add pieces of straight pipe and two bent pipes (which are called the long bent pipe and the short bent pipe) with different axial lengths to the compressor inlet; the diameters of these pipes and the impeller inlet are uniform, and the curvature radius of the centerline for these pipes is 80.5 mm. The distances be-tween the outlet of the two bent pipe and the impeller inlet are 1 and 1/4 times the diameter, respectively. For the convenience of the description in this paper, we use Case 0, Case 1 and Case 2 to refer to the models con-nected with the straight pipe, the long bent pipe and the short bent pipe.

The numerical method adopted in this paper is pre-sented in Ref. [8]. A spatially discrete grid of the cascade passage in the model is generated by using the IGG/ Autogrid software package. A flow field is plotted out and meshed in a structured mesh corresponding to the H-O type of block. Actually, the mesh type in the extension section and the diffuser is an H-type block, whereas the blade rows’ mesh grid is an O-type block. The grids of the inlet pipe and the volute are created manually using the IGG software package and are determined based on butterfly grid technology. The geometries and grid struc-tures of the two bent pipes are the same, and the centri-fugal compressors with the three different pipes are intro-duced in Ref. [7]. Furthermore, multigrid methods were employed to speed up the convergence to the solution.

For the numerical calculation, the commercial soft-ware named FINE/TURBO was utilized to solve the three-dimensional Reynolds-averaged Navier-Stokes equation in rotating and/or stationary frames of reference. In the present study, the Spalart-Allmaras model was used. The system of governing equations was advanced in time using an explicit Runge-Kutta scheme and dual time stepping. The unsteady calculation was set at the design point and near the stall point. The number of physical time steps is 120, and 50 pseudo-time steps were involved in one physical time step.

Experiment procedures

For the purpose of acquiring the test data, all the ex-periments and measurements for the centrifugal com-pressor with three different inlets were performed on a turbocharger test rig at the Turbocharging Technology Lab of the Beijing Institute of Technology. The com-pressor rotated with a turbine, which was driven by flow gas from an air supply. Fig. 1 shows an overview of the compressor inlet configuration with the straight and bend pipes for the experiment.

During the experiment, the test data for the compres-sor performance, which includes the mass flow, the rota-tional speed, the temperature at the inlet/outlet and the

Wang Leilei et al. Change of Inlet Geometry of Centrifugal Compressor Stage and its Influence on Compressor Performance 199

pressure, were measured. Among these test data, the mass flow was obtained indirectly by using the air flow meter with a double folium curve and calibration.

Table 1 Principal 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

Fig. 1 Compressor inlet configuration for the experiment. Dynamic pressure sensors provided information about

the dynamic pressure inside the compressor. The XCE- 062 sensors were adopted for the present experiment and were produced by the Kulite Corporation with the lowest response frequency of 175 kHz. It is difficult to test the unsteady pressure inside the impeller cascade passages because of the small structure size of the compressor and the high rotational speed of the impeller. Hence, the dy-namic pressure sensors must be installed in one side of the compressor shroud: the mid-chord position of the impeller was selected for the sensor installation in the experiment. This sensor may help us to discern the un-steady pressure fluctuation signal in order to analyze the flow field characteristics.

Fig. 2 shows the dynamic pressure sensors installed at key positions inside the centrifugal compressor, including

Fig. 2 The positions of the dynamic pressure sensors

the impeller inlet, the tip, the impeller outlet and the downstream stage. A detailed description of the installa-tion is given in Table 2. The radial positions downstream of the impeller take the radius R of the impeller outlet as a reference, and the reference position to the circumfer-ential angle is shown in Fig. 3. Confined by the volute, Sensor 1 and Sensor 2 are situated at 90° and 180°, re-spectively. The locations of Sensor 8 and Sensor 9 corre-spond to the maximum radius of 1.2R and thus are the immediate radial positions to the volute inlet. Table 2 Detailed description of these sensors’ positions

Serial number

Circumferential angle Axial or radial location

① 90°

② 180° Upstream main blade leading

edge, 3mm

③ 45° Tip of main blade, 50% chord

④ 150°

⑤ 165° R(impeller outlet)

⑥ 90°

⑦ 120° 1.1R(close to diffuser inlet)

⑧ 45°

⑨ 100° 1.1R(close to diffuser outlet)

Fig. 3 The reference position to the circumferential angle for the compressor volute

Data processing

The total pressure and static pressure are transformed to a dimensionless form with two formulae as follows.

p0 = p0′/ pref (1) p = p′/ pref (2)

where p0′ and p′ represent the values of the instantaneous total pressure and static pressure, respectively

The unsteady periodic change of the flow field in the compressor leads to a periodic unsteady fluctuation of the static pressure on the blade surface. A new coefficient Cpun is defined to reflect the fluctuation intensity ex-

200 J. Therm. Sci., Vol.22, No.3, 2013

pressed in formula (3) and is called the static pressure instability coefficient.

Cpun = pmax - pmin (3) where pmax and pmin are the maximum and minimum static pressure on the blade surface, respectively.

Let us express the time-averaged static pressure by using p as follows.

N

1

1i

ip p

N == ∑ (4)

where i is the ordinal number of the physical time step and N is the total number of physical time steps.

Results and analysis

The impacts on the compressor performance caused by different inlet pipe geometry

A centrifugal compressor inlet with a bent pipe may lead to radial and circumferential distortions. The con-clusion derived from the experimental study is that the flow distortion from the bent inlet mainly influences the impeller performance by causing non-uniform flow angle distortions at the inducer and a non-uniform total pres-sure distribution in the blade passages; as a result, the compressor performance dropped significantly[9]. When the compressor is connected with different inlet pipes, the change of the inlet pipe geometries lead to morphological differences in the inlet flow field.

Fig. 4 shows the compressor performance curves ob-tained under steady-state performance experiments with different inlet pipes. The diagram shows that the com-pressor pressure ratio and efficiency decreased when us-ing bent pipes compared to the straight inlet, and the compressor performance curve falls along the mass flow rates more obviously in case 1. The compressor per-formance curve in case 2 is between the two curves caused by case 0 and case 1.

When the compressor rotates at speeds of 80000 r

min-1 and 60000 r min−1, a comparison of the perform-ance curves shows that a larger performance amplitude distinction appears with higher revolution. This com-parison also indicates that the flow distortion induced by bent inlet pipes has a more sophisticated effect on the compressor performance as the rotational speed increase, whereas at the same rotational speed, the distortions have less influence on the compressor performance under a small mass flow rate, and the difference of the perform-ance curves are within the experimental error. The devia-tion between the performance curves becomes larger un-der large mass flow rate. These findings indicate that the inlet pipe geometry predominately influences the com-pressor performance under a large mass flow rate and that the performance decreases more obviously as the mass flow grows.

The inlet non-uniform flow caused by a bent pipe has an inevitable influence on the unsteady flow in the com-pressor and results in a forced response on the blade. Dickmann[10,11,12] researched centrifugal compressor blade vibrations produced by inlet pipe distortions, and the experiments showed that the blade vibration ampli-tude varies with mass flow changes. Figs. 5 and 6 present the static pressure pulsation spectrum at the impeller tip from dynamic pressure Sensor 3 mounted at the blade mid-chord position of the operating point near the choke condition (143% design point) at speeds of 80000 r min-1 and 60000 r min–1. The two plots show that the pressure fluctuation is predominated by the BPF (14th harmonic base frequency), and this frequency’s multiples and the amplitude increase as the rotation speed grows.

As shown in Fig. 5, compared to that of case 0, the amplitude of the BPF increases slightly in case 1 at high rotation speeds, whereas this frequency decreases in case 2 at the same rotation speed. For the bent inlet, the am-plitude of both the base frequency and the 28th harmonic base frequency increase; moreover, the latter increases significantly. The blade vibration is strengthened when

Fig. 4 Performance map for the compressors with the different inlets in this experiment


the increasing amplitude of the high frequency compo-nent is close or conforms to its natural frequency, which may cause the blade to produce high cycle fatigue.

At lower rotation speeds, Fig. 6 shows that the struc-ture of the amplitude spectrum for the bent inlet models shows no obvious distinction. The amplitude of the BPF is slightly higher with the bent inlet than with the straight inlet; however, the amplitude of the 28th harmonic base frequency decreases, and the changes of amplitude are approximately similar in the two types of bent inlet. This plot shows that the two types of bent inlet have similar and lower effects on the compressor performance com-pared to the straight inlet at lower rotation speeds.

Analysis of the variation of the total pressure in the inlet pipe

A bent inlet intake to a compressor would cause an inlet flow distortion, which is influenced by the down-

stream rotating impeller. A clear distinction can be ob-served in the production and development processes of the flow distortion compared with the case in which no downstream impeller exists. Because the results men-tioned above indicate that the inlet pipe geometry pre-dominately influences the compressor performance at high speeds and large mass flow rates, the following re-sults are discussed only for 80000 r min-1 and an opera-tion point near the stall point depending on the results of the numerical calculation.

The locations of the radial sections for analysis in the inlet pipe are shown in Fig. 7. For the two bent pipes, taking the inlet as a starting point, sections 1~4 are in similar positions as the curved section. Section 6 and section 7 have similar positions based on the impeller’s leading edge as a reference, and section 7 borders the reference surface. Because the axial distance is longer between the outlet of case 1 and the reference surface, a

Fig. 5 Frequency spectrum of the measured static pressure in the impeller tip at 80000 r min−1

Fig. 6 Frequency spectrum of the measured static pressure in the impeller tip at 60000 r min−1

Fig. 7 Reference sections inside the inlet pipe of the three models

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radial section 5 is added here (this section does not exist in case 2) to easily discover the orderliness of the varia-tion presented by the total pressure in the section. For the purpose of comparing the bent pipe with the straight pipe, section 5~7 are selected based on the reference location of the impeller’s leading edge as discussed earlier.

The reference sections are arranged along the pipe centerline in the downstream direction as shown in Fig. 8.

When an ideal fluid passes through a bend with a uni-form energy distribution, the static pressure increases with the increasing radius of the bend to balance the cen-trifugal forces toward the outer wall with a corresponding velocity decrease. Conversely, the static pressure lowered at the inner wall with a corresponding increase in veloc-ity. Influenced by the above condition, the total pressure distribution of the flow field is changed during the flow’s passage through the bent pipe, which would induce a lower pressure area close to the inner wall. This area corresponds to the total pressure distortion (Fig. 8). From section 5 to section 7, the figure suggests that the total pressure distortion continues to develop in the straight pipe connecting the bent pipe outlet with the leading edge of the impeller. As shown in Fig. 8, the discrepan-cies of the total pressure distortions caused by the three types of inlet pipes indicate that the distortion of the compressor inlet is related not only to the inlet pipe ge-ometry but also to the distance between the bent pipe outlet and the leading edge of the impeller and is also affected by the impeller rotation.

The structure of the total pressure distortion changes constantly with the air flow inside the bent pipe. As al-ready shown in Fig. 8, the total pressure distortion near the inner wall decreases significantly from section 3 to section 4. On section 4 of case 1, the distortion angle corresponding to the zone of the total pressure distortion is approximately 60 deg., and the radial position is ap-proximately 1/3 of the radius from the outside to the in-side. In addition, the total pressure drops more signifi-

cantly as the proximity of the distortion zone to the pipe wall increases. In contrast，on the same section, the dis-tortion angle of the total pressure distortion zone pro-duced by the short bent pipe is approximately 65 deg., and the radial position remains similar to the former case. The distribution of the total pressure distortions changes significantly on those sections between section 4 and section 7 for the three types of inlet pipes. At section 7, the intake flow of the straight pipe is only affected by the downstream impeller rotation for case 0, and the total pressure shows the distribution according to whether the value of the total pressure is high or low. For the interac-tion of the bent inlet and the downstream impeller rota-tion, the total pressure distribution near the outside wall in the bent pipe is similar to the phenomenon induced by the straight pipe, whereas the zone of the distortion in-creases slightly as the zone of the lowest total pressure decreases. On the same section, the radial position of the distortion zone reaches approximately 2/3 of the radius from the outside to the inside, which includes approxi-mately the whole spanwise location, and the distortion core is located at approximately 50% of the span. How-ever, in case 2, the radial position of the distortion zone remains approximately invariant, and the distortion core is located at approximately 90% of the span.

Analysis of the static pressure at 90% of the span for the main blade

Because of the total pressure distortion existing in the leading edge of the impeller, the periodicity of the non-uniform flow field in the compressor which impacts the progress of the impeller rotation, and the differences in the aerodynamic parameters would certainly lead to a redistribution of the static pressure around the blade sur-face.

(1) The static pressure distribution on the main blade surface

Near the choke condition of the centrifugal compres-sor, the static pressure distribution of the time-space dia-

Fig. 8 The distribution of the total pressure on all reference sections


gram for 90% of the span at the pressure side and the suction side of the main blade are shown in Fig. 9 and Fig. 10, in which subgraph a, subgraph b and subgraph c represent the cases of the three pipes connected with the compressor inlet.

Using the three inlet pipes, the static pressure distribu-tion on the pressure side (Fig. 9) shows that the pressure disturbance propagates upstream along the pressure blade surface (the orientation of the imaginary lines in the sub-graphs) due to the choke caused by the downstream compressor volute when the rotating impeller reaches the position of the compressor volute tongue. The pressure disturbance propagation arriving to the leading edge of the blade changes the lower static pressure distribution on the whole blade chord.

The fluctuations of the static pressure in case 0 are primarily impacted by the choke produced by the down-stream volute tongue as discussed earlier. However, the distortion caused by a bent inlet forms a new disturbance of the static pressure near the leading edge of the blade (locations marked with the dotted line circles in subgraph 9b and subgraph 9c). This new disturbance clearly propagates downstream along the blade surface at sonic speeds, and the propagation affects the distribution of the static pressure in the range of 0~0.6 chord; in this range, the area of the lower static pressure and the fluctuation amplitude of the static pressure would also increase.

The inlet flow distortions developing downstream also impact the static pressure distribution near the leading edge of the blade. Subgraph 9b and subgraph 9c are sig-nificantly different from subgraph 9a in the range of 0.65~1 chord, which means that the inlet flow distortion is diffused during the downstream propagation of the distortion. This characteristic changes the static pressure distribution and fluctuation in the area of the impeller outlet. Thus, the relative circumferential position be-tween the distorted region of the static pressure and the blocking effect from the volute tongue may influence the compressor performance.

Compared to case 1, the distortion caused by case 2 has a great effect on the lower static pressure area at the leading edge of the blade but has little effect on the static pressure distribution near the trailing edge of the blade. The differences should therefore be related to the extent of the distortion and the position of the distortion core before the leading edge of the blade when adopting the three different inlet pipes. The extent of the distortion induced by case 1 is larger, and the distortion core is lo-cated at approximately midspan, which may help the distortion transmit and produce a distinct impact on the trailing edge of the blade. In case 2, the extent of the dis-tortion is smaller, and the distortion core is located at approximately 90% of the span, which correspond to the location studied in this paper, and the distortion would be interposed by the tip leakage of the blade. Hence, case 2 is different from case 1. The distortion caused by case 2 has a dramatic effect on the leading edge of the blade; however, this distortion has little effect on the trailing edge of the blade.

As indicated in Fig. 10, the pressure disturbance also propagates upstream along the suction side in a manner similar to that observed along the pressure side; com-pared with case 0, the static pressure distribution near the leading edge of the blade acutely changes (locations marked with the dotted line circles at the leading edge in subgraph 10b and subgraph 10c) in the case of the mod-els with bent inlets. The fluctuation amplitude in this range clearly increases, which is similar to the pressure side. The inlet pipe of case 2 has a greater effect on the static pressure distribution and fluctuation than the inlet pipe of case 1 because the locations of the core area of the inlet distortion induced by the two bent pipes is dif-ferent. The distortion propagating downstream along the suction side significantly affects the area of higher static pressure near the trailing edge of the blade (locations marked with the dotted line circles at the trailing edge in subgraph 10b and subgraph 10c), and the fluctuation am-plitude of the static pressure also increases. The effects

Fig. 9 The distribution of the static pressure at 90% of the span on the pressure side of the main blade

204 J. Therm. Sci., Vol.22, No.3, 2013

Fig. 10 The distribution of the static pressure at 90% of the span on the suction side of the main blade

caused by the two bent inlets on the static pressure in the blade of the trailing edge are similar because the higher static pressure is largely distributed in the range of 0.9~1 chord near the trailing edge of the blade; moreover, the height of the blade is smaller in this range, so the differ-ence of the extent of the distortion and the position of the distortion core before the leading edge of the blade is subtle and slight.

(2) The fluctuation intensity coefficient for the surface of the main blade

Due to the inlet distortion, the distribution of the static pressure on the surface of the blade shows another state of fluctuation in the circumferential direction, which would cause additional fluctuations of the static pressure and have an inevitable impact on the reliability and ser-vice life of the blade.

Fig. 11 illustrates the variation of the fluctuation in-tensity coefficients at 90% of the span on the pressure side for the main blade for the three models. As shown in the figure, in the range of 0~0.15 chord of the leading edge of the blade, the values of the fluctuation intensity coefficients of case 1 and case 2 are slightly higher than that of case 0, and the distinction is more obvious be-tween case 2 and case 0. In the range of 0.22~0.7 chord of the blade, the change trends of the fluctuation intensity coefficient along the chord of the blade are appro-ximately uniform for the three types of inlet pipes, whereas the value with the bent inlets is approximately 20% higher than the value with the straight inlet, which can cause the fluctuation amplitude of the static pressure to increase. In the range of 0.7~0.85 chord of the blade, the fluctuation intensity coefficients are different in the three cases but not obviously. In the range of 0.85~1 chord of the blade, the fluctuation intensity coefficients with the bent inlets change more clearly, increasing ini-tially before decreasing. Both values of the two coeffi-cients are far larger than the value with the straight inlet. Fig. 12 illustrates the fluctuation intensity coefficients at 90% of the span on the suction side of the main blade for the three models. The two figures suggest that in the

range of 0~0.22 chord of the blade, the value of the fluc-tuation intensity coefficients for the bent inlets change dramatically and increase remarkably compared to those for the straight inlet. In the range of 0.22～0.7 chord of the blade, the variation trends of the fluctuation intensity coefficient along the chord of the blade are approxi-mately uniform for the three types of inlet pipes, whereas the values for the bent inlets are approximately 20% lower than the value for the straight inlet, which can de-crease the fluctuation amplitude of the static pressure. In the range of 0.7~0.8 chord of the blade, the fluctuation intensity coefficients are similar for the three cases. In the range of 0.85~1 chord of the blade, the values of the fluctuation intensity coefficients for the two bent inlets are higher, and the difference between these values is smaller.

Compared with the straight inlet, we attempt to ana-lyze the change of the fluctuation intensity coefficients plotted in Fig. 11 and Fig. 12. During the upstream de-velopment of the inlet distortion along the surface of the blade from the leading edge to the trailing edge, an ob-vious additive effect appears between the distortion and the static pressure on the surface of the blade, which would increase or decrease the fluctuation amplitude of

Fig. 11 The discrepancy of the fluctuation amplitude of the pressure side


Fig. 12 The discrepancy of the fluctuation amplitude of the

suction side

the static pressure on the blade surface of the relevant chord. The discrepancy of the fluctuation amplitude in-duced by the bent inlets can be explained by the static pressure distribution of the time-space diagram on the surfaces of the blade, the size of the inlet distortion area, the position of the distortion core, the change of the chord and the height of the blade.

(3) The distribution of the time-averaged static pres-sure on the surface of the blade

The results of the experiment as shown in Fig. 4 sug-gest that using the bent inlet leads to a decline of the effi-ciency of the compressor, and the change of blade load-ing crucially affects the efficiency. Zemp[13] performed an unsteady calculation of the inlet distortion for a centrifu-gal impeller with a diameter of 400 mm. The fluctuation amplitude of blade loading is thought to be maximized at the leading edge of the blade, and the fluctuation ampli-tude decreases as the mass flow rate decreases.

The change of the blade loading induced by the inlet distortion can be confirmed from the distribution of the time-averaged static pressure on the surface of the blade. Fig. 13 shows the distribution of the time-averaged static pressure at 90% span on the surface of the blade. If oper-ated near the choke condition of the compressor, the pressure gradient is greater along the flow direction in the range of 0.8~1 chord near the trailing edge. In the range of 0~0.2 chord near the leading edge, the values of the time-averaged static pressure on the pressure side and the suction side with the bent inlets are lower than the value with the straight inlet. Moreover, the decreasing under case 1 is more obvious, whereas the corresponding differences are barely changed. The changes may be formed by the flow angles, which are affected by the inlet distortion on the leading edge of the blade.

In the range of 0~0.8 chord, compared with case 0, the values of the time-averaged static pressure on the pres-sure side with the bent inlets decreased, and this value is lower for case 1. However, the values of the time-averaged static pressure on the suction side are similar in the

same range. This phenomenon indicates that the inlet distortion induced by the bent inlets decreases the ability of work in the intermediate chord of the blade. This phe-nomenon is the principle cause of the lower efficiency of the compressor at a large mass flow rate, and a decrease in efficiency is more obvious in case 1.

Fig. 13 The time-averaged static pressure on the surface of the blade

(4) The analysis of the static pressure fluctuation in

the diffuser The inlet distortion of the compressor changes the

structure of the blade flow field and induces effects downstream of the impeller. The generation of a “jet wake” structure at the outlet of a centrifugal impeller distorts the inlet flow of the diffuser and strongly affects the diffuser flow field[14], which has a profound effect on the compressor performance.

In this section, the influences on the static pressure fluctuation in the diffuser produced by using different inlet pipes are analyzed based on the experiment data. Fig. 14 shows the static pressure pulsation spectrum from dynamic pressure Sensor 6 mounted at the circumferen-tial position of 90 deg of the impeller under the choke condition of the compressor. Based on case 0, we find that the BPF and the 7th harmonic base frequency de-crease slightly for case 1, whereas the amplitude of the high frequency component, the 28th harmonic base fre-quency, increases. Compared with this case, the ampli-tude of the BPF increases slightly for case 2, whereas the amplitude of the high frequency component, the 28th harmonic base frequency, decreases. It is suggested that an obvious distinction exists at the inlet flow field of the diffuser in terms of the intensity of the unsteady pulsa-tion for the distortions induced by the two bent inlets.

Fig. 15 shows the average FFT distribution, computed within a circle at the inlet of the diffuser, of the 7th and 14th harmonic base frequencies, which plays a crucial role in the fluctuation of the static pressure. The main blade’s amplitude distribution of the 7th harmonic base

206 J. Therm. Sci., Vol.22, No.3, 2013

Fig. 14 The frequency spectrum of the measured static pressure at 90 deg of the diffuser inlet

Fig. 15 The distribution of the BPF within a circle at the diffuser inlet frequency is greater for case 0; an inverse result is ob-served for the bent inlets, and the values decrease near and below 0 deg.. For the three inlet pipes, the main blade’s amplitudes of the 14th harmonic base frequency change obviously, and the greater amplitude appears at the position of the tongue region (approximately 45 deg.) for the bent inlets.

Fig. 16 indicates the static pressure pulsation spectrum from dynamic pressure Sensor 8 mounted at the circum-ferential position of 45 deg according to the tongue region near the diffuser outlet. This figure clearly shows that the intensity of the fluctuation of the static pressure decreases after the airflow passes the diffuser. The blade rotating frequency of the upstream impeller is also

changed and affects the spectrum characteristics of the diffuser outlet. Compared with case 0, the amplitudes of the disturbance frequency induced by the blade BPF and the 7th harmonic base frequency under case 1 are approximately unchanged, whereas the amplitude of disturbance frequency induced by the 7th harmonic base frequency increases slightly. For case 2, the amplitudes of the disturbance frequency induced by the blade BPF and the 14th harmonic base frequency decrease slightly, whereas the amplitude of the disturbance frequency induced by the 7th harmonic base frequency clearly increases.

The BPF distribution of the compressor blade at the circumferential position is shown in Fig. 17. As with the diffuser inlet, for the three inlet pipes, the main blade’s

Fig. 16 The frequency spectrum of the measured static pressure at 45 deg. of the diffuser outlet


Fig. 17 The distribution of the BPF within a circle at the diffuser outlet amplitudes of the 7th harmonic base frequency at the circumference of the diffuser change with the position of the volute tongue region, at which location the greater amplitude of the 14th harmonic base frequency appears. The result demonstrates that the BPF distribution of the compressor blade passing the diffuser is clearly influenced by the volute tongue. The BPF distribution of the compressor blade with the straight inlet is different from the case with the bent inlets. The bent inlet can change the structure of the blade flow field, which greatly affects the flow field within the downstream vaneless diffuser.

Results

(1) The bent inlet results in a declining pressure ratio and efficiency of the centrifugal compressor, and the level of this decrease grows larger as the speed and flow rate increase. The distortion of the compressor inlet is related not only to the inlet pipe geometry but also to the distance between the bent pipe outlet and the leading edge of impeller and is also affected by the impeller rota-tion.

(2) The static pressure pulsation spectrum at the im-peller tip of the blade at the mid-chord position shows that the bent inlet impacts the blade flow field. At high speeds, the bent inlet greatly affects the amplitude-fre-quency characteristics, which are different for the long bent pipe and the short bent pipe.

(3) The inlet distortion of the compressor and the blocking effect from the volute tongue interact, resulting in a coupling effect that influences the fluctuation of the static pressure on the surface of the blade. The relative circumferential position between the distorted region and the blocking effect from the volute tongue may influence the compressor performance.

(4) The inlet distortion induced by a bent pipe changes the distribution of the static pressure and the fluctuation amplitude on the surface of the blade, and the degree of this influence is related to the extent of the distortion, the

position of the distortion core, the variation of the chord and the height of the blade. At large mass flow rates, the bent inlet changes the distribution of the blade loading on the surface of the blade and decreases the ability to work in the intermediate chord of the blade, which may lead to a lower efficiency; the long bent pipe has a more obvious influence.

(5) The inlet distortion changes the structure of the blade flow field of the compressor. It changes the inlet/ outlet amplitude-frequency characteristics of the diffuser and the BPF distribution of the diffuserr blade at the circumferential position, which impact the compressor performance.

Acknowledgements

This research was sponsored by the National Natural Science Foundation of China (No.51276017) and Ph.D. Programs Foundation of Ministry of Education of China (No. 20111101130002).

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The change of the inlet geometry of a centrifugal compressor stage and its influence on the compressor performance