4.1. Multiple Parameter Optimization
The impeller is a crucial component of the centrifugal compressor and is the only work component in the compressor. Whether the impeller structure is reasonable directly affects the working performance of the whole machine [
29,
30]. In this section, multi-parameter optimization design [
31,
32] will be carried out for the impeller structure to improve the aerodynamic performance of the centrifugal air compressor.
The efficiency of the centrifugal compressor was set as the optimization objective; four parameters—namely, the blade number, the blade angle at the inlet, the blade angle at the outlet and the blade wrap angle—were selected as the optimization parameters. The initial selections of the number of blades, inlet blade angle, outlet blade angle and wrap angle were 12, 50°, 60° and 50°, respectively.
The orthogonal method was adopted to conduct the optimization design according to the standard orthogonal test table L9 (3
4), in which three levels are set for each parameter. According to our research experience, the regions of blade number, blade angle at inlet, blade angle at outlet and wrap angle were set as 9–15, 40–60, 45–75 and 40–60, respectively.
Table 3 shows the orthogonal parameter and level. A, B, C and D represent four factors, and 1, 2 and 3 represent three levels.
According to the orthogonal table, nine individuals were determined and then nine compressors were designed, meshed and simulated.
Table 4 shows the details of parameter, pressure ratio and efficiency of the compressor. The minimum and maximum pressure ratios are 1.435 and 1.58, respectively. The minimum and maximum efficiencies are 77.74% and 83.12%, respectively.
In the orthogonal method, in order to evaluate the influence of every specific parameter, the average value
k of pressure ratio is defined as follows:
where
i is the level of the parameter,
n is the individual number of the corresponding parameter at level
i and
εj is the pressure ratio of tested compressors with level
i. The average value of efficiency is defined as the same formula.
Range
R describes the influence weight of each parameter, and can be defined as follows:
Table 5 shows the range analysis for compressor efficiency. According to
Table 5, the number of blades had the greatest impact on the efficiency of the centrifugal compressor, followed by the blade angle at the inlet. The blade angle at the outlet and the blade wrap angle also had an impact on the efficiency, but the impact was relatively weak. According to the analysis results, A3B1C1D3 was selected as the optimal result to design the finial optimal impeller of the compressor.
Table 6 shows the comparison of structural parameters between the baseline impeller and the optimized impeller.
Figure 10 shows the comparison of external characteristic curves of the baseline compressor and the optimal compressor. It can be seen from the figure that the change trend of external characteristics was basically the same for the baseline compressor and the optimal compressor, and the external characteristics of the optimal centrifugal compressor were significantly better than those of the baseline. Compared with the baseline, the pressure ratio and efficiency of the optimal compressor were improved, which indicates that the optimal blade geometry and flow field were more reasonable. At a design flow rate of 0.1 kg/s, the pressure ratios of the baseline compressor and the optimal compressor were 1.518 and 1.535, respectively, improving by 1.12%. Additionally, the efficiencies of the baseline compressor and the optimal compressor were 81.3% and 83.8%, respectively, improving by 2.50%.
Figure 11 and
Figure 12 show the velocity distribution of the baseline and the optimal centrifugal compressors at different blade height sections. It can be seen from the figure that the velocity in the flow field gradually increased with the increase in blade height due to the centrifugal force. Compared with the baseline, the local high-speed zone of the flow passage between the blades of the optimized centrifugal compressor was significantly reduced. At 70% and 90% of the blade height, the improvement of the local high-speed zone was the most obvious. The local high-speed zone of the optimized impeller blade was significantly reduced, and the flow field changed more evenly, which indicates that the matching degree of the impeller and volute was improved, and the flow separation has been effectively suppressed.
Figure 13 and
Figure 14 show the pressure distribution nephogram of the baseline and the optimized centrifugal compressor at different blade height sections. It can be seen that after the air entered from the impeller inlet, the pressure gradually increased as the impeller rotated. After impeller optimization, with the reduction of the local high-speed zone between blades, the local low-pressure zone in the flow field was also significantly reduced. The transition from the low-pressure zone at the impeller inlet to the high-pressure zone at the impeller outlet was more natural, the pressure distribution in the flow field was more uniform, the blade structure was more reasonable and the overall change of the flow field was more stable.
4.2. Pressure Fluctuation Suppression
Pressure fluctuation is a common phenomenon in a compressor, and severe pressure fluctuation will cause vibration, unstable operation and irregular noise, affecting the operation efficiency and stability of the compressor [
33,
34,
35]. For the centrifugal compressor, it is particularly important to maintain stable operation. Frequency domain analysis is a commonly used analysis method for pressure fluctuation. In frequency domain analysis, the time domain signal of pressure fluctuation is directly converted into frequency domain by the Fast Fourier Transform. In this work, the dimensionless pressure fluctuation coefficient
CP was used to represent the pressure fluctuation of the monitoring point, and it can be expressed as follows:
where
P is the pressure of the monitoring point,
is the average value of pressure,
is the air density and
is the circumferential speed at the impeller outlet.
Figure 15 shows the frequency domain of pressure fluctuation for the baseline compressor.
Figure 16 shows the frequency domain of pressure fluctuation for the optimal compressor. It can be seen that the peak of pressure fluctuation amplitude in the impeller occurred at the impeller rotation frequency (833 Hz) and its harmonic frequency. Among these monitoring points, the dominant frequencies for points M1, M2 and M3 were four times that of the rotation frequency, and the dominant frequencies for points M4 and M5 were the rotation frequency. On the whole, the amplitude of pressure fluctuation increased from the impeller inlet to the outlet. The pressure fluctuations at the monitoring points V1–V7 in volute were affected by the impeller rotation frequency and the blade passing frequency. The peak value of pressure fluctuation occurred at the impeller rotation frequency and blade passing frequency. The blade passing frequency was related to the rotation frequency and blade number. In the volute, the peak value of pressure fluctuation near the volute tongue was obviously large, and the stator-rotor interaction along the volute circumference gradually decreased, which made the peak of pressure fluctuation also gradually decrease.
The comparison of frequency domain between the baseline compressor and the optimal compressor shows that the pressure fluctuation of the optimal compressor was greatly reduced, especially near the volute tongue. The reason is that the optimal impeller geometry was more reasonable for the flow field in the impeller, and the pressure fluctuation was optimized and reduced. Based on the flow-field analysis, it can be seen that the airflow velocity at the inlet of the impeller increased, and the flow separation phenomenon at the outlet was suppressed. The local high-speed area in the flow channel shrank, and the low-pressure area correspondingly decreased. The flow field was more fluent, and the pressure gradient was more uniform, which resulted in a decrease in the intensity of pressure fluctuation. The peaks of the pressure fluctuation coefficient at point M1 in the baseline compressor and the optimal compressor were 0.0113 and 0.0073, respectively, decreasing by 35.1%. The peaks of the pressure fluctuation coefficient at point V1 in the baseline compressor and the optimal compressor were 0.0140 and 0.0041, respectively, decreasing by 70.5%.