# Study on Noise Model of an Automotive Axial Fan Based on Aerodynamic Load Force
## 4 Calculation Simulation and Verification of Automotive Axial Fan
### Noise Verification of Automotive Axial Fan
#### 4.2.2 Experimental Settings and Analysis Results of Automotive Axial Fan Noise  
1. Experimental Settings According to the structure parameters of the mathematical and simulation models, consistent experimental samples are created and the fan bench noise is verified. As shown  
Figure 10: Sound sensitivity analysis of different mesh sizes.  
Figure 9: The graph of sound pressure varying with time at receiving points.  
in Figure 11, the single fan is fixed to the bench using elastic rope in the four corners of the upper and lower beams, which ensures the vibration produced by the fan will not spread through the elastic rope to the bench and avoid noise interference. An external voltage-regulated power supply is connected to drive the fan to run under the specified speed. An LMS noise and vibration testing instrument is used for testing and verification.  
(2)Experimental analysis results  
According to the above simulation and experimental settings, the time-domain experimental results for the fan noise can be obtained combined with the corresponding boundary conditions. As shown in Figure 12, seven obvious noise peaks appear in one cycle, corresponding to the running noise of the fan's seven blades.  
The noise spectrum results can be obtained using an FFT transformation of the fan simulation and experimental noise time-domain calculation results that are compared with the numerical calculation results, as shown in Figure 13. It can be seen in Figure 13 that there is a small difference between the peak values of the numerical calculation results, simulation results, and experimental results in the middle- and low-frequency bands. Because the attenuation characteristics of high-frequency noise caused by damping in the transmission process are not considered in the numerical calculation, their results of a high frequency are obviously higher than those of the simulation and experiment.  
In order to further compare the differences between the numerical calculation results of the discrete noise and the simulation and experimental results, the first six orders of pulsating noise of the fan are extracted in this paper from Figure 13 (harmonic frequencies are 280 Hz, 560 Hz, 840 Hz, 1120 Hz, 1400 Hz, and 1680 Hz, respectively) for comparative analysis, as shown in Figure 14. The maximum differences between the first and third order pulsating noises through the numerical calculation and the simulation and experimental results are 0.6 dB(A) and 2.5 dB(A), respectively. The maximum differences between the fourth and sixth order pulsating noises through the numerical calculation and the simulation and experimental results are 2.2 dB(A) and 4.5 dB(A), respectively. The above comparison results do not exceed the error target requirement of 5 dB(A) of the spectrum peak, which indicates that the theoretical calculation model is reliable and the error is mainly caused by the simplification of the theoretical model to a certain extent, the inherent errors of the simulation analysis method, and the measurement errors in the experimental process.  
It can be seen from the above comparison results that the accuracy of the mathematical model of fan noise obtained from the load force using the new method proposed in this paper meets the requirements, which can be used as an important reference for the study of a high-precision noise mathematical model of low Mach-number fan and blade parameter optimizations.  
Figure 13: Single fan noise comparison of numerical calculation, simulation, and test results.