Analysis of vehicle vibration transfer characteristics based on flexible vehicle system and OTPA method
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摘要: 为了准确分析轨道车辆在较宽频域范围内的振动特性及传递规律，提出了一种基于弹性车辆系统动力学仿真模型的工况传递路径分析(OTPA)方法；建立了包含柔性轮对、构架和车体的弹性车辆系统动力学模型和与之结构参数完全相同的刚体模型，从时域的角度研究了轮对、构架和车体的振动特性，并将仿真结果与实测数据进行了对比，探究了弹性处理方式对车辆振动的影响，得出了振动能量的衰减规律；从频域的角度研究了在实测钢轨垂向不平顺的激励下，弹性车辆系统的振动特性；运用OTPA方法仿真分析了钢轨垂向不平顺结合车轮多边形的复杂工况下，车辆系统从轮对到构架至车体这一自下而上的振动传递过程当中垂向振动的主要传递路径。研究结果表明：车辆系统的弹性处理方式对整车振动有重要影响，弹性模型的轮对、构架和车体的振动加速度相比于刚体模型在中低频范围内更接近实测值，轴箱、构架和车体的最大振动幅值分别为250~450、30~40、3~4 m·s-2，由轮对至构架到车体，振动幅值呈一个数量级衰减；弹性模型的平稳性指标大于刚体模型，并且速度越大趋势越明显，车辆的弹性振动对运行性能的影响随着速度的提高而增大；车辆系统在复杂工况下，振动主要通过一系钢弹簧传递至构架，再通过空气弹簧和牵引拉杆传递至车内地板。Abstract: In order to accurately analyze the vibration characteristics and transfer rules of rail vehicles in the wide frequency domain, a scheme for operational transfer path analysis (OTPA) based on a dynamics simulation model of a flexible vehicle system was proposed. A dynamics model of a flexible vehicle system including flexible wheelsets, frame, and vehicle body, as well as a rigid body model with identical structural parameters was established. The vibration characteristics of the wheelset, frame, and vehicle body were studied from the perspective of time domain. The simulation results were compared with the measured data to explore the effect of flexible treatment on vehicle vibration. The attenuation law of vibration energy was obtained. Meanwhile, the vibration characteristics of the flexible vehicle system under the excitation of measured vertical irregularities of steel rails were investigated from the perspective of the frequency domain. The OTPA method was used to simulate and analyze the main transfer path of vertical vibration in the bottom-up vibration transfer process of the vehicle system from wheelset to frame and vehicle body under complex condition of vertical irregularities of steel rails combined with wheel polygon. Research results indicate that the flexible treatment method of the vehicle system has a significant impact on vehicle vibration. Compared with the rigid body model, the vibration accelerations of wheelsets, frame, and vehicle body in the flexible vehicle system model are closer to measured values in the mid-to-low frequency range. The maximum vibration amplitudes of axle box, frame, and vehicle body are 250-450, 30-40, and 3-4 m·s-2, respectively. The vibration amplitude attenuates by an order of magnitude from the wheelsets to the frame and vehicle body. The sperling index of the flexible vehicle system model is greater than that of the rigid body model, and the trend becomes more obvious as the speed increases. The impact of a vehicle's flexible vibration on the operational performance increases with the increase in speed. Under complex working conditions, the vibration of the vehicle system is mainly transferred to the frame through a series of steel springs, and then transferred to the interior floor through air springs and traction rods.
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Key words:
- rail transit /
- flexible vehicle system /
- OPTA /
- vertical irregularity of steel rail /
- wheel polygon /
- rigid body model /
- transfer path
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图 14 车体中心横向振动加速度曲线
Figure 14. Curves of transverse vibration acceleration in center of vehicle body
图 15 车体中心垂向振动加速度曲线
Figure 15. Curves of vertical vibration acceleration in center of vehicle body
图 16 车体前端横向平稳性曲线
Figure 16. Curves of transverse smoothness of front end of vehicle body
图 17 车体前端垂向平稳性曲线
Figure 17. Curves of vertical smoothness of front end of vehicle body
图 19 轮对垂向振动功率谱密度曲线
Figure 19. Curve of vertical vibration power spectral density of wheelset
图 20 空簧座垂向振动功率谱密度曲线
Figure 20. Curve of vertical vibration power spectral density of air spring seat
图 21 车体垂向振动功率谱密度曲线
Figure 21. Curve of vertical vibration power spectral density of vehicle body
图 23 构架钢簧帽垂向振动加速度曲线
Figure 23. Curves of vertical vibration acceleration of frame steel spring cap
图 24 构架空簧座垂向振动加速度曲线
Figure 24. Curves of vertical vibration acceleration of frame air spring seat
图 25 车体前端垂向振动加速度曲线
Figure 25. Curves of vertical vibration acceleration of front end of vehicle body
表 1 模态对比
Table 1. Comparison of modals
部件 阶次 频率/Hz 误差率/% 自由模态 正则化后模态 车体 1 8.000 8.130 1.60 2 8.345 8.669 3.74 3 9.687 9.863 1.78 构架 1 42.070 42.636 1.33 2 68.943 69.863 1.32 3 86.183 87.215 1.18 轮对 1 97.190 97.837 0.66 2 102.380 102.353 0.03 3 102.380 102.353 0.03 
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表 2 振动加速度对比
Table 2. Comparison of vibration accelerations
m·s-2 振动加速度 刚体模型 弹性模型 实测值 均方根值 最大值 均方根值 最大值 均方根值 最大值 轮对横向 2.08 14.05 1.72 12.87 2.17 13.17 轮对垂向 4.29 32.54 5.66 69.85 7.19 69.51 构架横向 1.32 5.48 1.98 14.86 1.85 12.76 构架垂向 2.32 8.99 2.84 26.15 3.15 19.26 车体横向 0.11 0.43 0.21 0.82 0.13 0.43 车体垂向 0.23 0.82 0.35 1.33 0.37 1.31
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表 3 OTPA仿真工况
Table 3. OTPA simulation conditions
轨道激励 车轮多边形阶数 磨耗深度/mm 时间/s 德国低干扰谱 0 2.0 1、3、5 0.10 6.0 7、9、11、13 0.05 8.0 15、17、19 0.03 6.0 21、23、25 0.01 6.0 钢轨垂向不平顺 0 0.8 1、3、5 0.10 2.4 7、9、11、13 0.05 3.2 15、17、19 0.03 2.4 21、23、25 0.01 2.4
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表 4 车体垂向振动不同峰值频率下的贡献量
Table 4. Vertical vibration contribution amounts of vehicle body at different peak frequencies
dB 位置 贡献量 2.63 Hz 28.69 Hz 48.54 Hz 80.58 Hz 钢簧帽垂向 89.88 86.08 73.76 77.33 钢簧帽横向 99.73 92.07 83.40 90.36 抗蛇行减振器构架端垂向 119.91 101.77 82.67 96.07 抗蛇行减振器构架端横向 117.34 112.36 93.95 98.28 空簧座垂向 122.84 103.91 94.61 91.21 空簧座横向 126.72 105.83 89.11 104.14 牵引拉杆构架端横向 112.59 95.38 93.54 84.79 牵引拉杆构架端垂向 126.51 110.28 94.40 82.94
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