频谱分析在转子动平衡中的应用#中英文翻译#外文翻译匹配

2009 届本科毕业论文（设计） 相关中英文翻译资料 资料 题目： 频谱分析在转子动平衡中的应用 学生姓名：徐军锋 所在院系：机电学院 所学专业：机电技术教育 指导老师：付素芳 APPLICATION OF FREQUENCY SPECTRUM ANALYSIS IN THE ROTATOR MOVING EQUILIBRIUM ABSTRACT The experimental equipment is developed to simulate the rotator vibration. The running state of machine is simulated by using different running conditions. The vibration caused by non-equilibrium mass is analyzed and discussed for first order with focus load. The effective method is found out by using frequency spectrum analysis. INTRODUCTION In the conventional island of nuclear power plant, turbine generator set is a very important equipment in which the core thermal energy is transferred into electric energy. When the turbine generator set has run for long time, the original equilibrium of system would be upset because of the remnant deformed of the metal, abrasion or damaging of the components etc. As a result, the vibration will be increased. So it is necessary to adjust the equilibrium at spot. On the other hand, a large turbine generator set also needs the work for adjusting the equilibrium in the process of manufacture, debugging, installation and operation. The moving equilibrium technique at spot is an important means to eliminate the violent vibration of the turbine generator set. We could have a definitely view for the vibration type, vibration power source and vibration property by analyzing the vibration of the turbine generator set or doing some special experiment. When the vibration signal is obtained, the frequency spectrum could be used to analysis the vibration signal in order to diagnose quickly. Using frequency spectrum analysis, the electrical signal of vibration that is obtained by the vibration sensor and has a wide frequency range will be decomposed into several main frequency compositions. Different frequency compositions have different influence on the turbine generator set. The frequency spectrum analysis is a very useful method to study the vibration of turbine generator set. The vibration caused by rotator mass non-equilibrium with concentrated load is discussed and analyzed in this paper. And an effective method to prevent the vibration is presented by using frequency spectrum analysis. 1 EXPERIMENTAL EQUIPMENT The experimental system consists of the motor, shaft coupling and rotor etc. Its structure is very simple. The rotor is driven by the motor directly. Its rotating speed could be adjusted in a wide range. The system could be operated smoothly and reliably. The rated current of the motor is 2.5 A, the output power is 250 W. The field excitation of the motor is provided by the 220 V AC power source which is commutated by the speed regulator, the armature current of the motor is also provided by the same power source. It is adjusted by the compressor governor. Through adjusting the output voltage of compressor governor, the motor could be of step-less speed regulated at the range 0 10000 r/min, the rate of velocity increasing could be 800 r/min. The length of the experimental equipment is 1200 mm, the width is 108 mm, the mass is about 45 kg, the diameter of the shaft is 9.5 mm, the length of the shaft is 500 mm and the maximum deflection is less than 0.005 0.015 mm. Any position along the axial direction could be selected as experimental abutment point. The diameter and mass of the experimental rotating table is 7619 mm and 600 g, respectively. The arrangement of experimental equipment is shown in Fig.1. Fig. 1 The arrangement of experimental equipment The electrical vortex sensors are used to measure the relatively displacement or vibration for axis to bearing pedestal. They are installed in the x and y directions at the sensor support, respectively. They do not touch the shaft, and could be used to directly measure the vibration signal of the rotation shaft. The flashing phase-measurer is used to measure the rotator speed and the phase of the shaft. 2 THE FREQUENCY SPECTRUM ANALYSIS OF VIBRATION SIGNAL The real vibration of turbine generator set is the most of simple harmonic periodic motion. Its wave type is also made of many simple harmonic motion. In order to analysis the vibration, we should study the wave type, the frequency composition of the vibration, and the amplitudes. Frequency can be used as x-axis to describe the vibration in the frequency-domain. The method decomposing the vibration into its various frequency components in frequency components in the frequency domain is called frequency spectrum analysis. The purpose of frequency spectrum analysis is to decompose the signals into different compositions. So the vibration becomes the simple harmonic motion including different amplitudes, frequencies and phases. In the frequency spectrum of rotator vibration, the different frequency is often corresponding to the different reason. If we can find the frequency composition of the vibration signal, the reason of vibration will be discovered. There are about 80% accidents caused by rotator non-equilibrium in the vibration accidents of turbine generator set happened in the spot, and 90% accidents caused by rotator mass non-equilibrium. In the experiments described in this paper we study the vibration caused by rotator mass non-equilibrium by using the method of frequency spectrum analysis and the influence coefficient method of finding equilibrium to determine the position of rotator mass non-equilibrium. 3 EXPERIMENTAL RESULTS AND ANALYSIS The Bode diagram shown as Fig. 2 is about the horizontal vibration characteristics of rotator. It shows that the critical velocity of rotator is about 2605 r/min, the maximum amplitude of rotator corresponding to the bearing shell is 371 m, the amplitude of select frequency (AMP-1X) is 360 m, and the phase difference (PHA-1X) is -36. The mark position in Fig. 3 is the maximum value of vibration, and also is the position where the phase angle changes. When the rotator speed is smaller than 2152 r/min, the relative phase angle is -130. When the rotator speed increases to 2605 r/min, the amplitude will increase to the maximum value. That is to say, the vibration amplitude increases rapidly in the rotator speed range 2152 2605 r/min. Fig. 2 The Bode diagram of rotator horizontal vibration characteristic The phase change )94)130(36( is slightly greater than 90. According to the eccentric forced vibration theory for single free dimension, the case in the Fig.2 is caused by resonance. The maximum amplitude occurs at the position where the phase change is slightly greater than 90, because of the damping. In this case, the angle frequency of rotator speed equals to that of exciting force, that is, 1 . The angle frequency of rotator speed can be considered as first order critical rotator speed. The critical rotator speed is related to material of rotator, geometrical shape, size, structure, and supporting conditions etc. It is the inherent characteristics of the rotator system, and it is not related to the external conditions. The influence factors are mainly temperature and supporting rigidity. The experiments have been done under the condition of constant temperature. So supporting rigidity is the main influence factor. The experiments have been done by using the method of adding the non-equilibrium mass. The location adding non-equilibrium mass is determined by influence coefficient method that finding equilibrium. The non-equilibrium mass can change the system stiffness. But, according to m, when the m is very small, 1 will almost not change. If mm increases, 1 will be decreased. The first order critical rotator speed of the system will be decreased. These can be obtained from the experiments. The frequency spectrum figures show that the amplitude is obviously large when the frequency is one times (1X). It is monotonously increasing before the speed reaches critical rotator speed. The cases of frequency two times (2X), three times (3X) and four times (4X) all exist, but these amplitudes are very small and can be neglected. It is impossible that rotating table or system axis appear crosswise cracking when the experimental velocity is not high enough. The main reason causing vibration is that the mass of rotating table around axis is not uniform. Fig. 3 and Fig. 4 are the frequency spectrum diagrams that obtained through adding the different non-equilibrium mass to rotator system, respectively. Fig. 3 shows that the first order critical rotator speed is basically not changed, and is 2312 r/min, but the amplitude of frequency decreases. The horizontal amplitude of one times (1X) decreases to 145 m, and the vertical amplitude decreases to 134 m. For Fig. 4, they are 116 m and 87 m, respectively. The horizontal amplitude of 1X in Fig. 3 is decreases to 145 m, the vertical amplitude is decreased from 360 m to 134 m. The result that the second adding non-equilibrium mass is shown in Fig. 4. From this case, we can see: the 1X component is very obvious in the frequency spectrum. There are components for 2X, 3X, 4X, but their amplitudes are very small. They are not the main components of vibration. The 1X amplitude changes very small when change the system stiffness. This is determined by the location of adding non-equilibrium mass. The vibration amplitude can be effectively controlled only by calculation to find out the non-equilibrium point and non-equilibrium mass, then adding the same equilibrium mass at its opposite direction. Fig. 3 The frequence spectrum diagram of the first adding non-equilibrium mass Fig. 4 The frequence spectrum diagram of the second adding non-equilibrium mass 4 CONCLUSION (1) In the frequency spectrum figure, the one times frequency 1X components is too large. When the malfunction about bearing pedestal stiffness and axis joint join defect is not considered. The reason why the vibration is greater is the rotator non-equilibrium mass. (2) The one times frequency 1X amplitude is decreased by changing system stiffness. The decreasing amplitude is determined by the location of adding non-equilibrium mass. (3) The location of non-equilibrium mass is determined by the influence coefficient method. It is needed to find the non-equilibrium point and non-equilibrium mass by calculation. Then add the same equilibrium mass at the opposite direction. (4) The adding non-equilibrium mass is so small that it can not cause the large change of the system first order critical velocity. 频谱分析在转子动平衡中的应用 摘 要 在模拟旋转机械振动的实验装置上，通过不同的选择来模拟机器的运行状态，对单跨集中载荷情况下转子由于不平衡质量引起的振动进行了分析和讨论，并用频谱分析的方法找到了有效的解决办法。 介绍 在传统的岛屿核电站中， 汽 轮发电机组是一种非常重要 的核 热能转换成电能的设备。当 汽轮 发电机组 经长时间运转后 ,原来的 系统 平衡会因 金属的 残余变形、部件的 磨损或损坏 而遭到破环 。结果 ,系统的机械 振动将会 因此 增加。 所以 因此有必要 进行现场平衡 调整。另一方面 ,一个大型汽轮发电机组 在制作工序中 也需要调整平衡、调试、安装和运行。现场动平衡技术是消除 汽轮发电机组 剧烈振动的 一种重要的手段 。 我们可以 通过对汽轮发电机组的振动进行分析或做一些特殊的实验 明确 了解振动的类型、振动动力源和振动特性 。 当 获得 振动信号 之后， 频谱可以用来分析振动信号，以便 迅速 诊断。利用频谱分析 由振动感应器获得的 电机 的振动 信号，并 将广泛的频率范围 分解为几个主要的频率成分。不同频率成分 对汽轮发电机组有着不同的影响。频谱分析是 研究汽轮发电机组振动的一个非常有用的方法 。 本文将对 由于 集中载荷引起的转子质量不平衡 进行 讨论和分析 ， 并且给出了一个 利 用 频谱分析 有效 防止振动 的方法。 1 实验设备 实验系统由电机、联轴器及转子等 构成 ， 它的结构是非常简单的。转子 由 电机直接驱动 ， 它的转速可 进行大范围 调节。该系统可以顺利 、可靠的运作 。电机的额定电流为 2.5 A,输出功率是 250 W。 电动机 的外部 励磁 由 220 V 交流电源 经 调节器整流后 提供 的 ,发动机的 电枢电流也 是 相同 的电源 提供 的 。它是由压缩机 调节器进行 调整 控制 ， 通过调整压缩机输出电压， 电动机 的速度 可逐步减少调节 至 范围 0 10000 转 /分 ， 速度递增可达 800 r /分。 实验设备的长度是 1200 毫米 ,宽度为 108 毫米 ,质量是大约 45 公斤 , 轴的 直径是 9.5 毫米 ， 轴的 长度 500 毫米 ,最大挠度小于 0.005 0.015 毫米。 可以选择沿轴向的 任何位置作为实验的支承点。实验 用 转 盘的直径和质量分别 为 7619毫米 ,重 600 克。实验设备的 安装如 图 1 所示 。 图 1 实验设备 的安装 电涡流传感器是用来测量 中轴 相对 轴承底座的 位移或振动的。 传感器 分别 被安装在 X 和 Y 方向 提供信号传递 。 它 们不接触 轴 ,但 可以直接用于测量 转动轴的振动信号。 闪动相位测量仪 是用来测量 转子 速度 和传动轴的相位 。 2 振动 信号 频谱分析 轮机发电机组 真正的振动的 是 最简单的谐波周期运动 。它的波型也是由许多简单的谐波运动 构成 。为了分析振动 ,我们应该学习 振动的 的波型、频率 组成 和振幅。 在频域中 频率 可以被看 作 X 轴来 描述振动。 在频域中将 振动 分解 成 各种频率成分的方法 叫做频谱分析。频谱分析的目的是为了 将振动信号 分解成不同的信号成分。所以振动 被分解 成为谐波运动包括不同的振幅、频率和阶段。 在转子振动的频谱 中 ,不同的频率通常是对应于不同的原因。如果我们能找到振动信号的 频率成分 ，也 就 会发现引起 振动 的原因 。 在轮机发电机组工作现场 有大约 80%的 振动 事故 是由转子不平衡 引起的 ,90%的事故 是 由 转子质量不平衡 引起的 。在本文中所描述的实验 中 我们会采用频谱分析的方法 研究转子质量平衡引起的振动 和 用 影响系数法 找出转子质量不平衡 的位置 。 3 实验结果及分析 图 2所示伯德 图 是关于转子的 横向振动特性。 图示转子的 临界速度 约是 2605 r /分钟 ,转子相对应轴承壳 最大限度的 振幅为 371m,选择频率 (AMP-1X)是360m、相位差 (PHA-1X)是 -36。 在图 2 中 标记位置是最大 幅度的振动 ,也就是振动相位变化的位置。当转子速度小于 2152 r /分钟 ,相对相位是