中英文翻译--频谱分析在转子动平衡中的应用-精品.pdf

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1、.-本科毕业论文（设计）本科毕业论文（设计）相关中英文翻译资料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资料题目：频谱分析在转子动平衡中的应用学生：学生：所在院系：机电学院所在院系：机电学院所学专业：机电技术教育所学专业：机电技术教育指导老师：指导老师：.可修编.-APPLICATION OF FREQUENCY SPECTRUM ANALYSIS IN THE ROTATOR MOVINGAPPLICATION OF FREQUENCY SPECTRUM ANALYSIS IN THE ROTATOR MOVINGEQUILIBRIUMEQUILIBRIUMABSTRACTAB

2、STRACTThe experimental equipment is developed to simulate the rotator vibration.Therunning state of machine is simulated by using different running conditions.Thevibration caused by non-equilibrium mass is analyzed and discussed for first order withfocus load.The effective method is found out by usi

3、ng frequency spectrum analysis.INTRODUCTIONINTRODUCTIONIn the conventional island of nuclear power plant,turbine generator set is a veryimportant equipment in which the core thermal energy is transferred into electric energy.When the turbine generator set has run for long time,the original equilibri

4、um of systemwould be upset because of the remnant deformed of the metal,abrasion or damaging ofthe ponents etc.As a result,the vibration will be increased.So it is necessary to adjustthe equilibrium at spot.On the other hand,a large turbine generator set also needs thework for adjusting the equilibr

5、ium in the process of manufacture,debugging,installation and operation.The moving equilibrium technique at spot is an importantmeans to eliminate the violent vibration of the turbine generator set.We could have a definitely view for the vibration type,vibration power source andvibration property by

6、analyzing the vibration of the turbine generator set or doing somespecial experiment.When the vibration signal is obtained,the frequency spectrum couldbe used to analysis the vibration signal in order to diagnose quickly.Using frequencyspectrum analysis,the electrical signal of vibration that is obt

7、ained by the vibrationsensor and has a wide frequency range will be deposed into several main frequencypositions.Different frequency positions have different influence on the turbine generatorset.The frequency spectrum analysis is a very useful method to study the vibration ofturbine generator set.T

8、he vibration caused by rotator mass non-equilibrium with concentrated load isdiscussed and analyzed in this paper.And an effective method to prevent the vibration ispresented by using frequency spectrum analysis.1 EXPERIMENTALEQUIPMENT1 EXPERIMENTALEQUIPMENTThe experimental system consists of the mo

9、tor,shaft coupling and rotor etc.Itsstructure is very simple.The rotor is driven by the motor directly.Its rotating speedcould 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

10、motor is provided by the 220 V AC power source which.可修编.-is mutated by the speed regulator,the armature current of the motor is also provided bythe same power source.It is adjusted by the pressor governor.Through adjusting theoutput voltage of pressor governor,the motor could be of step-less speed

11、regulated atthe range 010000 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,themass is about 45 kg,the diameter of the shaft is 9.5 mm,the length of the shaft is 500mm and the maximum deflection is less than 0.0050.015

12、 mm.Any position along theaxial direction could be selected as experimental abutment point.The diameter andmass of the experimental rotating table is 7619 mm and 600 g,respectively.Thearrangement of experimental equipment is shown in Fig.1.Fig.1 The arrangement of experimental equipmentThe electrica

13、l vortex sensors are used to measure the relatively displacement orvibration for axis to bearing pedestal.They are installed in the x and y directions at thesensor support,respectively.They do not touch the shaft,and could be used to directlymeasure the vibration signal of the rotation shaft.The fla

14、shing phase-measurer is usedto measure the rotator speed and the phase of the shaft.2 THE FREQUENCY SPECTRUM ANALYSIS OF VIBRATION SIGNAL2 THE FREQUENCY SPECTRUM ANALYSIS OF VIBRATION SIGNALThe real vibration of turbine generator set is the most of simple harmonic periodic.可修编.-motion.Its wave type

15、is also made of many simple harmonic motion.In order toanalysis the vibration,we should study the wave type,the frequency position of thevibration,and the amplitudes.Frequency can be used as x-axis to describe the vibrationin the frequency-domain.The method deposing the vibration into its various fr

16、equencyponents in frequency ponents in the frequency domain is called frequency spectrumanalysis.The purpose of frequency spectrum analysis is to depose the signals intodifferent positions.So the vibration bees the simple harmonic motion including differentamplitudes,frequencies and phases.In the fr

17、equency spectrum of rotator vibration,the different frequency is oftencorresponding to the different reason.If we can find the frequency position of thevibration signal,the reason of vibration will be discovered.There are about 80%accidents caused by rotator non-equilibrium in the vibrationaccidents

18、 of turbine generator set happened in the spot,and 90%accidents caused byrotator mass non-equilibrium.In the experiments described in this paper we study thevibration caused by rotator mass non-equilibrium by using the method of frequencyspectrum analysis and the influence coefficient method of find

19、ing equilibrium todetermine the position of rotator mass non-equilibrium.3EXPERIMENTAL RESULTS AND ANALYSIS3EXPERIMENTAL RESULTS AND ANALYSISThe Bode diagram shown as Fig.2 is about the horizontal vibration characteristicsof rotator.It shows that the critical velocity of rotator is about 2605 r/min,

20、themaximum amplitude of rotator corresponding to the bearing shell is 371 m,theamplitude 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 theposition where the phase angle changes.When the rotator speed

21、 is smaller than 2152r/min,the relative phase angle is-130.When the rotator speed increases to 2605r/min,the amplitude will increase to the maximum value.That is to say,the vibrationamplitude increases rapidly in the rotator speed range 21522605 r/min.可修编.-Fig.2 The Bode diagram of rotator horizonta

22、l vibration characteristicThe phase change(36(130)94)is slightly greater than 90.According to the eccentric forced vibration theory for single free dimension,the case inthe Fig.2 is caused by resonance.The maximum amplitude occurs at the position wherethe phase change is slightly greater than 90,bec

23、ause of the damping.In this case,the angle frequency of rotator speed equals to that of exciting force,that is,1.Theangle 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,a

24、nd supporting conditions etc.It is the inherent characteristics of the rotatorsystem,and it is not related to the external conditions.The influence factors are mainlytemperature and supporting rigidity.The experiments have been done under the condition of constant temperature.Sosupporting rigidity i

25、s the main influence factor.The experiments have been done byusing the method of adding the non-equilibrium mass.The location addingnon-equilibrium mass is determined by influence coefficient method that findingequilibrium.The non-equilibrium mass can change the system stiffness.But,accordingto m,wh

26、en the m is very small,1will almost not change.If mmincreases,1will be decreased.The first order critical rotator speed of the system will bedecreased.These can be obtained from the experiments.The frequency spectrum figures show that the amplitude is obviously large whenthe frequency is one times(1

27、X).It is monotonously increasing before the speed reachescritical rotator speed.The cases of frequency two times(2X),three times(3X)and fourtimes(4X)all exist,but these amplitudes are very small and can be neglected.It isimpossible that rotating table or system axis appear crosswise cracking when th

28、e.可修编.-experimental velocity is not high enough.The main reason causing vibration is that themass of rotating table around axis is not uniform.Fig.3 and Fig.4 are the frequencyspectrum diagrams that obtained through adding the different non-equilibrium mass torotator system,respectively.Fig.3 shows

29、that the first order critical rotator speed isbasically not changed,and is 2312 r/min,but the amplitude of frequency decreases.Thehorizontal amplitude of one times(1X)decreases to 145 m,and the vertical amplitudedecreases to 134 m.For Fig.4,they are 116 m and 87 m,respectively.Thehorizontal amplitud

30、e of 1X in Fig.3 is decreases to 145m,the vertical amplitude isdecreased from 360 m to 134 m.The result that the second adding non-equilibrium mass is shown in Fig.4.Fromthis case,we can see:the 1X ponent is very obvious in the frequency spectrum.Thereare ponents for 2X,3X,4X,but their amplitudes ar

31、e very small.They are not the mainponents of vibration.The 1X amplitude changes very small when change the systemstiffness.This is determined by the location of adding non-equilibrium mass.The vibrationamplitudecanbeeffectivelycontrolledonlybycalculationtofindoutthenon-equilibrium point and non-equi

32、librium mass,then adding the same equilibriummass at its opposite direction.Fig.3 The frequence spectrum diagram of the first adding non-equilibrium mass.可修编.-Fig.4 The frequence spectrum diagram of the second adding non-equilibrium mass4 CONCLUSION4 CONCLUSION(1)In the frequency spectrum figure,the

33、 one times frequency 1X ponents is too large.When the malfunction about bearing pedestal stiffness and axis joint join defect is notconsidered.The reason why the vibration is greater is the rotator non-equilibriummass.(2)The one times frequency 1X amplitude is decreased by changing system stiffness.

34、The decreasing amplitude is determined by the location of adding non-equilibriummass.(3)The location of non-equilibrium mass is determined by the influence coefficientmethod.It is needed to find the non-equilibrium point and non-equilibrium mass bycalculation.Then add the same equilibrium mass at th

35、e opposite direction.(4)The adding non-equilibrium mass is so small that it can not cause the large changeof the system first order critical velocity.可修编.-频谱分析在转子动平衡中的应用频谱分析在转子动平衡中的应用摘要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。介绍介绍在传统的岛屿核电站中，汽轮发电机组是一种非常

36、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

37、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 1 实验设备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围 010000 转/分，速度递增可达 8

38、00r/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。.可修编.-图图 1 1 实验设备的安装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 2 振动信号频谱分析振

39、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

40、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 3 实验结果及分析实验结果及分析图 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 2605r/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 2152r/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,在旋转速度围 2152.可修编.-2605 r/分振幅迅速增加。图图 2 2 转子横向振动特性伯德

41、图转子横向振动特性伯德图相变(36(130)94)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,1。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 些可从实验中获知。频谱的数

42、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

43、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.可修编m,当质量变化非常小时的,1几乎没有改变。如果转子质量增加,1将减少。系统的一阶临界旋转速度将会下降。这.-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。图图 3 3 频率频谱首次加入不平衡质量图频率频谱首次加入不平衡质量图图图 4 4 频率频谱第二次加入不平衡质量图频率频谱第二次加入不平衡质量图4 4 结论结论（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.可修编.-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。.可修编