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1、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： 所在院系：机电学院所学专业：机电技术教育指导老师：APPLICATION OF FREQUENCY SPECTRUM ANALYSIS IN THE ROTATOR MOVING EQUILIBRIUMABSTRACTThe experimental equipment is developed to simulate the rotator vibration. The running state of machine is simulated by using different running con
2、ditions. 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
3、 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
4、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
5、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 fre
6、quency 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. Diff
7、erent 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 p
8、aper. 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
9、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 th
10、e 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 010000 r/min, the rate of velocity increasing could be 800 r/min. The length of the
11、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 0.0050.015 mm. Any position along the axial direction could be selected as experimental abutment point. The
12、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 equipmentThe electrical vortex sensors are used to measure the relatively displacement or vibration for axis to
13、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 sha
14、ft. 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 compositio
15、n 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
16、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 diffe
17、rent 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 mas
18、s 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 EXPE
19、RIMENTAL 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 frequ
20、ency (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 increas
21、es 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 21522605 r/min. Fig. 2 The Bode diagram of rotator horizontal vibration characteristicThe phase changeis slightly greater than 90. According to the
22、 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 ex
23、citing force, that is, . 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 sy
24、stem, 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 m
25、ethod 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 , when the m is very small, will almost not change. If mm increases,w
26、ill 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
27、 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. T
28、he 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 sp
29、eed 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 decr
30、eases 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 ve
31、ry 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
32、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 massFig. 4 The frequence spectrum diagram of the second adding non-equilibrium mass4 CONCLUSION (1) In the frequency spectrum fig
33、ure, 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
34、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 sa
35、me 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
36、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
37、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
38、 /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。
39、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
40、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-130。当转子速度增加到2605 r /分钟,振幅将增加到最大值。也就是说, 在旋转速度范围2152 2605 r /分内振幅迅速增加。图2 转子横向振动特性伯德图相变略大于90。根据单一自由维度的偏心受迫振动理论可知图2中的情况是由共振引起的。受阻尼影响最大的振幅发生位置的相变略大于90。在这种
41、情况下,转子速度旋转角频率等于激振力的角频率,也就是说,。转子旋转角频率速度可视为一阶临界转子速度。转子的旋转速度取决于转子的材质、几何形状、大小、结构和支撑条件等，与外部条件无关，影响因素主要是温度和支撑刚度。该实验在恒温条件下完成，因此支承刚度是主要影响因素。这个实验通过添加非平衡质量完成。添加非平衡质量的位置取决于用影响系数法寻找平衡。不平衡质量可以改变系统刚度。但是,根据,当质量变化非常小时的,几乎没有改变。如果转子质量增加,将减少。系统的一阶临界旋转速度将会下降。这些可从实验中获知。频谱的数字显示频率为1倍时振幅明显很大。在旋转速度达到临界速度之前它是单调增加的。对所有平衡情况的频率
42、两倍、三倍、四倍时都存在，但这些振幅的存在很小,可以忽略不计。当试验速度远不够高的情况下系统轴出现横向裂缝裂的情况是不可能的。引起振动主要的原因是绕轴旋转的转盘不统一。图3、4的频谱图是通过添加不同的转子质量使系统分别达到平衡。图3显示了一阶临界转子速度基本上没有改变,为 2312 r /分,但振幅频率下降。频率的一倍下降到145m、垂直振幅下降到134m。在图4中它们分别为116m和87m。如图3是横向振幅减少到145m、垂直振幅从360m下降至134m。第二次增加非平衡质量的结果如图4所示。在从这种情况下，我们可以看到：在频谱中1倍频率的组成部分是非常明显的。2倍， 3倍， 4倍频率的部分
43、也同样存在，但其幅度是非常小的。它们不是振动的主要组成部分。当更改系统刚度时，1倍频率的幅度变化是非常小的。这是由添加不平衡质量的位置所决定的。我们可以只通过计算来找出非均衡点和非平衡质量有效的控制振幅,然后在其相反的方向添加相同的平衡质量。图3 频率频谱首次加入不平衡质量图图4 频率频谱第二次加入不平衡质量图4 结论（1）在频谱图，一次频率一倍的部分太大。当有关轴承座刚度和轴心联合的障碍不考虑时，振动变大的原因是转子不平衡质量。（2）通过改变系统刚度降低了一倍次频率的振幅，下降幅度取决于平衡的位置。（3）不平衡质量的位置,是由影响系数法确定的。需要找到不平衡点和计算不平衡质量，然后将平衡质量加在相反的方向。（4）添加不平衡质量很小,所以它并不会造成系统一阶临界速度发生大的变化。设计巴巴工作室