Study on the Influence of Radial Inlet Chamber Splitter Blades on the Oblique Flow Compressor Performance

Abstract

The oblique flow compressor is one of the important components in the compressed air energy storage (CAES) system. The structural shape of the radial inlet chamber (RIC) directly affects the compressor performance, and a reasonable RIC design should achieve the smallest total pressure loss and outlet distortion as much as possible to meet the structural design. To study the influence of splitter blades, 4 RICs equipped with different numbers of splitter blades are designed, and the performance of 4 RICs and the overall performance of the compressor is calculated. The results show that with the increase in the number of splitter blades, the stall margin increases from 6.3% to 13.94%. At the design point, the isentropic efficiency is highest for the RIC with 17 splitter blades, and the pressure ratio is highest for the RIC with 11 splitter blades. Compared with the direct axial intake mode, the uniformity of the relative leakage distribution and the attack angle distribution of the impeller leading edge under 4 radial intake modes are poor. However, with an increase in the number of splitter blades, the uniformity of the relative tip leakage and the attack angle distribution gradually increase. The flow loss of RIC will increase simultaneously, though the uniformity of the outlet aerodynamic parameters distribution improves, and the influence on the downstream component performance gradually weakens. There is an optimal number of splitter blades in RIC, which balances the total pressure loss and distortion coefficient.

1. Introduction

Compressed air energy storage (CAES) system has the advantages of high efficiency, long operation life, and large energy storage capacity. This has important application value in energy utilization, such as peak regulation and frequency regulation of power grid, wind, and photovoltaic coupling complementarity [1,2,3]. The CAES system is considered one of the energy storage technologies with the most development potential. The compressor is one of the most important components of the compressed air energy storage system, and its performance directly determines the efficiency and economy of the energy storage system [4,5]. The radial inlet chamber (RIC) is one of the important static components in the compressor unit. Its function is to introduce airflow from the inlet pipe or heat exchanger into the impeller inlet. The pressure loss and distortion will occur after air flows through the RIC, which directly affects the work capacity and efficiency of the impeller. The structural shape of RIC has an important impact on compressor performance. The RIC and volute can reduce the efficiency of the centrifugal compressor stage by 1% to 4% [6]. A reasonable RIC design should meet the smallest possible flow loss and keep the outlet airflow as uniform as possible. Due to the complex geometry of the RIC, a systematic advanced design method has not yet been developed [7,8]. The simplified design is often used in engineering, which ensures the convergence of the entire flow channel in the RIC to limit the area of several key sections, and then determine the geometric parameters of each section. At present, the RIC study mainly concentrates on the internal flow and its influence on compressor performance.

Under the low flow condition, tip leakage vortex movement will occur at the blade tip, which will affect the impeller leading edge attack angle and thus affect the efficiency and stability of the impeller [9,10,11,12]. Galindo [13] studied the effects of different vortices at the impeller inlet and found that opposite vortices can improve the stall margin. Wang [14] calculated the internal flow field of the RIC in detail by numerical methods and found that the propagation of the vortex system generated by the airflow in the annular channel along the flow direction and the impeller rotation were the main reasons for the flow loss and outlet distortion. Flathers [15,16,17] calculated the RIC performance with and without splitter vanes through numerical simulation, measured the flow field of RIC by using a three-hole probe, and obtained the distribution of swirl angle, static pressure, and total pressure along the circumference of the outlet section. And the two results were in good agreement. Koch [18] used numerical simulation to calculate the performance of RIC with splitter vanes and found that the splitter vanes made the outlet airflow of the RIC more uniform and the outlet tangential air angle smaller, but increased the total pressure loss. Pazzi [19] analyzed the Mach number and yaw angle distribution of the RIC outlet and found a high-speed and low-pressure area above the RIC outlet. There are symmetrical large yaw angle areas at the left and right ends of the outlet section, caused by the propagation of two symmetrical vortices in the RIC spiral channel along the flow direction. Hohlweg [20] studied the influence of centrifugal compressor stage axial distance on the stage performance and found that the more compact RICs have greater outlet distortion and the stage efficiency is lower. Kim [21] first designed a RIC with splitter vanes and then designed a new splitter vane structure according to the flow field of the first type of RIC. The airflow in the new RIC was more reasonable, the pressure loss was small, and the outlet flow was more uniform. Han [22,23,24,25] calculated the flow field inside the RIC and its influence on centrifugal compressor performance by numerical calculation, and found that the flow loss caused by the RIC had little influence on compressor performance, but the outlet distortion was the main reason for the deterioration of the performance of downstream components. In addition, the aerodynamic parameters of each section inside RIC were measured in detail through experiments. The results showed that the sudden expansion of the spiral channel produced a vortex, which was transmitted to the outlet and distorted. Saladino [26] and Yagi [27] selected several key geometric parameters which have a great impact on the RIC and optimized them. The results proved that the proposed optimization method was credible, and the performance and stage efficiency of the RIC could be effectively improved after optimization. Chen [28] parametrically modeled the RIC annular convergence channel and established the surrogate model between the geometric parameters of the annular channel and RIC performance through the parallel neural network algorithm, and finally optimized by the genetic algorithm. The annular channel expanded, and the meridian profile was smoother after the optimization. The RIC total pressure loss coefficient was reduced by 16.45%, and the stage total pressure ratio increased by 0.6% at the designed point. In summary, the RIC optimization mainly focuses on its geometry and shape, and less involves the influence of the number of internal splitter blades.

In this paper, the influence of the number of splitter blades on RIC performance and downstream components is studied to guide the RIC design.

2. RIC Design

The RIC includes an intake channel, a spiral channel, and an annular convergence channel. Figure 1 shows the RIC geometry structure. The gas density is assumed to be constant during the design process because the flow rate change in the RIC is small. The inlet volume flow rate is calculated from the compressor design flow rate and gas density. The inlet flow rate is calculated from the inlet area and volume flow. The diameter of the RIC inlet round tube D₁ is calculated from the inlet flow conditions. The hub diameter D₃ and outlet diameter D₄ are given by the impeller inlet geometry. Based on the existing design experience [7], the convergence level σ₁ of the intake channel is 1.73. A streamlined support plate is set in the air intake channel to achieve structural requirements while minimizing interference to the internal flow field. The spiral channel is a symmetrical structure along the left and right planes of the middle line of the structure, and the flow area of each flow section is determined by Equation (5), showing a trend of gradual reduction. The convergence degree σ₂ of the annular convergence channel is set as 2.71. The annular convergence channel relative turning radius λ₁ and λ₂ are set as 1.24 and 0.17, respectively.

The parameters are defined as follows:

$${\sigma }_1=\frac{\pi D_1^2}{4bl}$$

(1)

$$q_{v-in}=\frac{m_d}{\rho }$$

(2)

$$c_{in}=\frac{4q_{v-in}}{\pi D_1^2}$$

(3)

$$c_{{180}^{°}}=c_{in}·{\sigma }_1$$

(4)

$$A_{\theta }=\frac{q_{v-in}}{c_{{180}^{°}}}·\frac{\theta }{{360}^{°}}$$

(5)

$${\sigma }_2=\frac{4b{D}_2}{D_4^2-D_3^2}$$

(6)

$${\lambda }_1=\frac{R_1}{b}$$

(7)

$${\lambda }_2=\frac{R_2}{b}$$

(8)

RICs with 7, 11, and 17 splitter blades are designed based on the flow characteristics of the RIC, as shown in Figure 2. The blade shape and distance are reasonably designed and distributed in the spiral channel according to the RIC internal flow field.

3. Computational Model and Numerical Methodology

3.1. Computational Model

To analyze the influence of the RIC with different splitter blades number on the oblique flow compressor performance, the selected calculation model includes the RIC, inlet guide vanes, impeller, and the vaneless diffuser, as shown in Figure 3. In comparison, the performance of the oblique flow compressor is calculated separately (model 0). There are five models for simulation calculation, as shown in

Table 1. 5 models with different RIC.

Model	Radial Inlet Chamber	Compressor
Model 0	-	IGV + Impeller + Vaneless diffuser
Model 1	0 blades
Model 2	7 blades
Model 3	11 blades
Model 4	17 blades

. The detailed designed parameters of the compressor are listed in

Table 2. The designed parameters of the compressor.

Parameters	Value
Rotation speed (r/min)	3000
Inlet total pressure (Pa)	99,330
Designed pressure ratio	1.9
Number of IGV	12
Number of impellers	13
Tip clearance (mm)	1.5
Impeller inlet width (mm)	1460
Impeller outlet width (mm)	115

.

3.2. Evaluation Indicators

To quantitatively evaluate the influence of the RIC on the oblique flow compressor performance, the performance of the RIC, oblique flow compressor, and the whole machine were defined, respectively. The total pressure loss coefficient K_p, outlet distortion coefficient $\zeta$, and outlet swirl angle $\theta$ are performance indicators for evaluating the RIC. The performance of the whole machine is reflected by the total pressure ratio $\epsilon$, isentropic efficiency $\eta$, surge margin SM, and stable operation range OM. The expressions of the above parameters are defined as follows:

Total pressure loss coefficient of RIC:

$$K_p=\frac{P_{in}^{*}-P_{out}^{*}}{P_{in}^{*}-P_{in}}$$

(9)

Outlet distortion coefficient:

$$\zeta =\frac{\sqrt{\frac{1}{n}{\displaystyle \sum _{i=1}^n{\left(P_i-{\overline{P}}_{out}\right)}^2}}}{{\overline{P}}_{out}}\times 100\%$$

(10)

In the formula, ${\overline{P}}_{out}$ is the average static pressure of all sampling points at the RIC outlet. $P_i$ is the static pressure of the sampling point i in the outlet section, and n is the number of sampling points. The smaller the distortion coefficient, the more uniform the pressure distribution.

The tangential airflow angle of the cross-section:

$$\theta =abs\left[\arctan (\frac{V_c}{V_m})\right]$$

(11)

In the formula, V_c is circumferential velocity and V_m is meridian velocity.

The total pressure ratio of the whole machine:

$$\epsilon =\frac{P_{out}^{*}}{P_{in}^{*}}$$

(12)

The isentropic efficiency of the whole machine:

$$\eta =\frac{{\left(\frac{P_{out}^{*}}{P_{in}^{*}}\right)}^{\frac{k-1}{k}}-1}{\left(\frac{T_{out}^{*}}{T_{in}^{*}}\right)-1}$$

(13)

The surge margin of the whole machine:

$$SM=\left[\frac{{\left(\frac{\epsilon }{m}\right)}_s}{{\left(\frac{\epsilon }{m}\right)}_d}-1\right]\times 100\%$$

(14)

The stable operation range of the whole machine:

$$OM=\left(\frac{m_c-m_s}{m_d}\right)\times 100\%$$

(15)

In the formula, T and T* are static temperature and total temperature, respectively. P and P* are static pressure and total pressure, respectively. k is the adiabatic coefficient. The subscript s represents the surge condition, c represents the chock condition, and d represents the designed condition.

3.3. Mesh Generation and Calculation Method

The calculation domains in this paper are the RIC, adjustable inlet guide vanes, semi-open oblique flow impeller, and vaneless diffuser (Figure 4). The RIC is meshed unstructured and locally encrypted in the annular convergence channel, support plate, and splitter blades. Hexahedral structured meshing is used for adjustable inlet guide vanes, impellers, and vaneless diffusers. The first boundary layer thickness of the entire computing domain is set to 2 × 10⁻⁶ m, the boundary layer is set to 10 layers, and most of the wall y+ values are less than 1. Considering the calculation accuracy and resource, the grid independence of the RIC and compressor were verified, respectively. The results are shown in Figure 5. The final grid number of each component is shown in

Table 3. Mesh number of compressor and RIC.

Section	Mesh Number/Million
Compressor	1.57
RIC without blade	1.98
RIC with 7 blades	2.32
RIC with 11 blades	2.75
RIC with 17 blades	3.17

.

In this paper, the full channel steady numerical simulation of the compressor model is carried out. The SST turbulence model is selected to solve the three-dimensional constant Navier-Stokes equations, and the discrete spatial format is selected for high resolution. The ideal gas calculates working fluid, and the solid wall adopts the adiabatic non-slip boundary condition. The dynamic and static interface adopts the frozen rotor method to transmit information. The inlet is given the total temperature to total pressure ratio boundary condition, and the turbulence intensity is set as 5%. Since the compressor outlet is sensitive to pressure when choking and sensitive to mass flow during stall conditions, the outlet is given a small back pressure at the beginning of the calculation, and the mass flow boundary condition is given at the outlet when calculated near the design point. During the calculation process, the relative mass flow error and residual of the inlet and outlet are monitored. The calculation is considered convergence when the relative error is less than 0.5%, and the residual is less than 1 × 10⁻⁵.

3.4. Verification of Numerical Calculation Method

In this paper, the HPCC centrifugal impeller with public geometric and experimental data [29,30] is selected to verify the numerical method accuracy. The HPCC impeller is a semi-open impeller with 15 main blades and 15 splitter blades. Its designed mass flow rate is 4.54 kg/s, designed rotation speed is 21,789 r/min, and designed pressure ratio is 4. The experimental and numerical results of the HPCC impeller are shown in Figure 6. It can be seen from the figure that the total pressure ratio and isentropic efficiency of the simulated value and experimental value are basically the same, and the maximum error is 0.8% and 1.2%, respectively. It indicates that the numerical method in this paper can accurately simulate the compressor’s overall performance, which can be used for further research and analysis.

4. Calculation Results and Discussion

4.1. Overall Performance

4.1.1. Whole Machine

The performance of the oblique flow compressor with the RIC equipped with different numbers of splitter blades is shown in Figure 7. It can be seen from the figure that the pressure ratio–mass flow characteristic curve and isentropic efficiency–mass flow characteristic curve trends of the five models are roughly the same. The pressure ratio and isentropic efficiency of model 0 are greater than those of the other four models in the entire operating range. The pressure ratio and isentropic efficiency of the other four models are not much different under the large mass flow condition. The pressure ratio of model 3 is larger than that of the other three models under the small mass flow condition. The isentropic efficiency of models 3 and 4 is basically equal and larger than that of the other two models in the whole operation range. Figure 8 shows the stable operation range, surge margin, pressure ratio, and isentropic efficiency at the design point for the five models. In Figure 8a, comparing model 1 to model 4, it can be seen that with the increase of the number of splitter blades in the RIC, the stable operation range and surge margin of the whole machine are gradually improved. The machine stable operation range is increased from 39.09% to 45.33%, and the surge margin is increased from 6.3% to 13.94%. Comparing model 4 and model 0, it can be seen that the difference between the stable operation range and surge margin is only 2.79% and 0.85%. In Figure 8b, comparing model 1 to model 4, it can be seen that the machine’s isentropic efficiency increases with the increase of the number of splitter blades at the design point. The isentropic efficiency of model 4 increases by 0.33% compared with model 1, and the isentropic efficiency of model 4 decreases by 1.49% compared with model 0. The machine pressure ratio first increases and then decreases with the increase of the splitter blade number in the RIC. The pressure ratio of model 3 is the largest at the design point, and the pressure ratio of model 3 is reduced by 1.14% compared to model 0.

4.1.2. RIC

Figure 9 shows the RIC performance of four models. It can be seen that the total pressure loss coefficient increases gradually with an increased number of splitter blades. In addition, the distortion coefficient first increases and then remains basically unchanged with the increase of splitter blades. The distortion coefficient of both models is around 2.14% when the number of splitter blades is 11 or 17. When the number of splitter blades increases to a certain number, the flow separation becomes more intense after the flow through the splitter blades, the mixing loss inside the RIC increases, and the outlet flow uniformity is less changed than before. From the above analysis, it can be concluded that there is an optimal number of splitter blades in the RIC so that the total pressure loss and distortion coefficient are minimized in balance.

The distribution of aerodynamic parameters at the RIC outlet directly determines the downstream components’ performance. The flow above and below the RIC outlet is quite different, so the aerodynamic parameters of the upper and lower straight lines are extracted for comparison. The aerodynamic parameter distribution of the two straight lines perpendicular to them is extracted at the same time. The four straight-line distributions are shown in Figure 10. Figure 11 shows the static pressure distribution along four straight lines for four models. It can be seen from the figure that the static pressure of the four models gradually decreases from hub to shroud in the four straight lines, and the static pressure distribution difference of the four models on the four straight lines is within 1000 Pa. Since the RIC is symmetrical to the left and right, the static pressure is distributed similarly on L3 and L4. In addition, the static pressure value of L2 is the largest overall, followed by L3 and L4, and L1 is the smallest. It can be seen from Figure 11a that the static pressure is arranged in L1 as model 3, model 4, model 2, and model 1. Compared to Figure 11b, the static pressure on L2 is arranged in the order of model 1, model 2, model 3, and model 4. The order of the above two is not completely opposite, which results in the flow loss caused by the increase of the RIC splitter blade number. Compared with the static pressure distribution of the four models on L3 and L4, the static pressure of the four models is arranged oppositely. The reason is the rotation of the impeller, which affects the upstream components and destroys the static pressure distribution uniformity at the RIC outlet.

Figure 12 shows the swirl angle distribution on four straight lines for four models. It can be seen that the swirl angle on the L1 and L2 of the four models are generally smaller than those of the L3 and L4. Comparing the swirl angle distribution of the four models on L1, it can be seen that the swirl angle of the four models on L1 firstly decreases from hub to shroud, then stabilizes, and then increases. Except for model 2, the L1 has a large swirl angle near the hub, and the swirl angle on the L1 of the other models is between 0 to 4.5°. The swirl angle on the L2 of the four models is within 4°, and the swirl angle of the four models shows continuous fluctuations from hub to shroud. Comparing the distribution of swirl angles on L3 and L4, it can be seen that the swirl angle of the four models on L3 and L4 are within 20°, which is much larger than on L1 and L2. The swirl angle distribution of the four models on L3 and L4 are basically the same. Except for the proximity to the hub and shroud regions, the swirl angle from hub to shroud shows a trend of first increasing and then decreasing, and the swirl angle is basically symmetrical in the interval. The swirl angle distribution on L3 and L4 is ranked from largest to smallest; model, model 2, model 3, and model 4, and the order is the same as the number of splitter blades.

4.1.3. Impeller

To facilitate the description of impeller circumferential asymmetry, the impeller blades are ordered, and the angle of the circumferential position is defined, as shown in Figure 13.

Figure 14 shows the circumferential change trend of the attack angle at a 50% span in the impeller leading edge of the five models. As can be seen from the figure, except for some blades, the impeller leading edge attack angle for the 5 models is basically between −6°–2°.

In model 0, the attack angle at 50% span of the impeller leading edge shows good periodicity along the circumferential distribution, and the attack angle is about −3.5°. After adding the RIC, the impeller leading edge attack angle is obviously destroyed along the circumferential periodic distribution. But with the increase of the number of splitter blades in RIC, the flow is more evenly distributed along the circumferential, and the attack angle distribution of the impeller leading edge tends to be uniform. Comparing the attack angle distribution of each blade of the five models, it can be seen that the attack angle of model 0 is greater than that of the other four models at the leading edge of blade No. 11-1. The attack angle of model 0 is smaller than that of other models at the leading edge of blades No. 3,4 and No. 8–10. On the other blades, the attack angle of model 0 and the other four models is similar. After flowing out of the RIC convergence channel, the flows mix under the RIC and then travel along the flow direction to the impeller leading edge, with the impeller rotation, making the attack angle of the No. 11-1 blade larger. Comparing the attack angle of the No. 1 blade of the five models, it can be seen that after adding the splitter blades in RIC, the attack angle of the No. 1 blade remains basically the same as the original model.

Figure 15 shows the relative airflow angle circumferential distribution at a 50% span of the impeller trailing edge for the five models. It can be seen from the figure that the relative airflow angle of the five models is distributed between 50° and 68°. The impeller tailing edge relative airflow angle in model 0 is distributed circumferentially around 60°. The uniformity of the relative airflow angle distribution of the other four models is poor, but with the increase in the number of splitter blades in RIC, the uniformity of the impeller outlet’s relative airflow angle distribution increases. Compared with model 0, the outlet relative airflow angle of the other 4 models are larger on the blade of No.13 and 1 but smaller on the blade of No. 2–5, 7, 9, and 10. It can be seen that the uneven distribution of the impeller leading edge attack angle will continue to propagate to the impeller trailing edge as the impeller rotates.

4.2. Tip Leakage

In this paper, the relative tip leakage at different mass flows in five models is numerically calculated (Figure 16). It can be seen from the figure that the relative tip leakage of the five models has similar trends with mass flow, which decreases with the mass flow increase, and the decline rate is significantly accelerated at large mass flow. Since the structure of the RIC is symmetrical, the airflow mix at the RIC convergence channel and impeller inlet flow rate are small due to the loss caused by the RIC itself. Therefore, the relative tip leakage of model 0 is greater than that of the other four models in the entire mass flow range, and the relative tip leakage of model 0 at the near surge point and near choke point is quite different from other models, but the difference in other working conditions is small. The relative tip leakage of the other four models varies slightly across the whole mass flow range.

Figure 17 shows the relative tip leakage of each impeller blade at the design point of the five models. Compared with model 0, the relative leakage distribution of impeller blades of the other four models is uneven. The relative leakage of the blades in these four models have a similar distribution trend along the circumferential direction, gradually rising from blade No.1 to blade No.5, then descending and flattening, and then continuing to rise from blade No. 10 and then beginning to decline at blade No.12. The relative tip leakage of the impeller at the blade No. 5 and No. 12 are large. The relative position of blade No. 5 is above the RIC outlet, and the flow velocity is high, so the leakage is large. Blade No. 12 is below the RIC outlet, where more airflow mixes, so the tip leakage is also large. In addition, with the increase of the RIC splitter blade number, the uniformity of the relative tip leakage shows a better trend.

4.3. Analysis of Flow Field

4.3.1. RIC

The number of splitter blades has an important influence on RIC internal flow, affecting downstream component performance. Figure 18 shows the static pressure distribution at a 50% span of the RIC support plate for four models. It can be seen that the static pressure distribution of the four models is basically axisymmetric in this section. There are symmetrical low-pressure regions on both sides below the support plate because of the flow separation at the support plate downstream. In model 1, the airflow mostly turns and flows out above the convergence channel. There is a high-speed low-pressure area above the convergence channel under the centrifugal force and a low-speed high-pressure area at the spiral channel bottom. With the increase of the splitter blades, the low-pressure area above the convergence channel and the high-pressure area at the spiral channel bottom gradually decrease, and the static pressure distribution is more evenly distributed along the circumferential direction. However, when the air flows through the splitter blades, it produces shock loss at the leading edge and flow separation at the tailing edge, which increases the flow loss inside the RIC. In addition, when the number of splitter blades is seven, the airflow produces severe flow separation and large vortices under the four splitter blades, which has a great impact on RIC performance.

Figure 19 shows the static pressure distribution and two-dimensional streamline distribution of the RIC outlet section in four models. Comparing the static pressure distribution of the four models, it can be seen that the static pressure distribution tends to be symmetrical at the entire interface because the RIC is structurally symmetrical. But due to the influence of downstream impeller rotation, the symmetry of the static pressure distribution is destroyed. There is a low-pressure area above the RIC outlet and a high-pressure area below it if there is no splitter blade. With the increase in the number of splitter blades, the high-pressure area and low-pressure area gradually decrease, and the static pressure distribution of the section tends to be uniform. Comparing the streamline diagrams of the four models, it can be seen that there is a symmetrical vortex structure in the RIC outlet in model 1, and as the number of splitter blades increases, these vortices begin to develop upward along the shroud and gradually weaken, and basically disappear when the splitter blade number reaches 17.

4.3.2. Impeller

The distribution of aerodynamic parameters of the impeller inlet directly affects the impeller’s workability. Figure 20 gives the distribution of the impeller inlet swirl angle. The swirl angle is distributed uniformly throughout the impeller inlet at around 0°. In the other four models, the swirl angle distribution of the impeller inlet was uneven. There are large swirl angle areas near the lower left and lower right of the hub. These two swirl angle regions gradually decrease as the splitter blade number increases, and these two regions basically disappear in model 4. Near the shroud, there are alternately large swirl angle regions, and these areas gradually decrease as the splitter blade number increases, but the area hardly changes when the splitter blade number reaches 11.

4.3.3. Blade to Blade

Figure 21 shows the entropy increase distribution along the flow channel of the compressor at 10% span, 50% span, and 90% span of the five models. Comparing the entropy increase distribution of different blade span in the five models, it can be seen that the uniformity of entropy increase is better in the circumferential distribution at the blade root, blade middle, and blade tip in model 0. At the 10% compressor span, the area with a large entropy increase is mainly in blade channels. There is a large entropy increase area at the blade tip due to the wake caused by the airflow in the impeller tailing edge. The mixing loss in the entire vaneless diffuser is large, so the entropy increase is also larger, and the entropy increase in other regions is small. In the model with RIC, the distribution of entropy increase in the entire flow channel basically becomes more uniform with the increase of splitter blade number. In model 4, the entropy increase is consistent with model 0 along the circumferential distribution throughout the flow channel. Due to the RIC influence before the inlet guide vanes, the airflow uniformity at the RIC outlet is poor. The mass flow of the airflow above the RIC outlet is large, the airflow propagates along the flow direction, and a larger range of entropy increase area is produced at the inlet guide vanes tip. The airflow produces a large flow separation between the No. 4 blade and the No. 7 blade, the airflow is more disordered, and a small entropy increase is produced near the suction surface. The airflow continues along the flow direction, with a greater entropy increase at the impeller tailing edge and vaneless diffuser. Similarly, the airflow mass flow is small below the RIC outlet near the hub, and the airflow propagates along the flow direction to produce a large entropy increase area at the blade root of the No. 12 blade and No. 3 blade.

5. Conclusions

In this paper, 4 RIC models with different numbers of splitter blades are designed, RIC performance and whole stage performance are calculated by numerical simulation and compared with the original model, and the following conclusions are obtained:

• With the increase of splitter blades in the RIC, the stable operation range and surge margin of the whole stage gradually increase, the pressure ratio increases first and then decreases, and the isentropic efficiency gradually increases at the design point. The surge margin increases from 6.3% to 13.94%. At the design point, the isentropic efficiency of the whole stage (RIC with 17 blades) increases by 0.33% compared to the 0 splitter blade (the lowest) and reduces by 1.49% compared to that without RIC. The total pressure ratio of the whole stage (RIC with 11 blades) is the largest, and the total pressure ratio is reduced by 1.14% compared to the compressor without RIC;
• The relative tip leakage of the models with different numbers of splitter blades have a similar trend with mass flow, and all decrease with the increase of mass flow. The decline rate accelerates significantly at large mass flow. Compared with the compressor without RIC, the relative tip leakage and attack angle distribution uniformity of the other 4 compressors with different RIC are poor. With the increase of the splitter blade number, the uniformity of the relative leakage and attack angle distribution shows a better trend;
• With the increase of the splitter blade number in RIC, the flow loss in RIC increases, but the uniformity of outlet aerodynamic parameters distribution is significantly improved. It is possible to determine the best compromise between total pressure loss and outlet distortion.

The specific number of RIC splitter blades should also consider the relationship to the number of impeller blades. The influence of compressors with different flow coefficients for different types needs further study.