支撑刚度对汽轮机组轴承座振动影响分析

近年来汽轮发电机组轴承座振动问题越来越多,而关于支撑刚度对汽机组轴承座振动影响的研究较少。建立大型汽轮机组柔性支撑下转子—轴承—支撑系统不平衡响应分析有限元模型,结合某大型汽轮机组实测振动数据,研究了支撑刚度对汽轮机组振动特性的影响。研究结果表明,支撑刚度对机组振动特性影响较大,在同样大小激励力作用下,柔性支撑时机组轴振较小,但是轴承座振动较大。     

菏泽电厂1号汽轮机组轴承振动诊断及处理

介绍了菏泽发电厂1号汽轮机组2号轴承振动异常的诊断过程及处理方法。从2002年10月2号轴承振动异常到2006年5月对转子加平衡块,经过了3年多不断地分析和处理,从增加轴承座动刚度着手,将2号轴振控制在报警值以下,基本上解决了2号轴承振动故障。     

某电厂1号机组振动故障诊断与处理

某电厂1号机组在运行过程中依次出现5号、4号轴承瓦振爬升及7号轴承相对轴振阶跃等故障现象.针对上述问题,进行多种振动测试,研究了汽轮机强迫振动幅值、激振力、支撑系统动刚度之间的关系,分析振动机理,并采取不同的消振措施,为解决同类机组的振动故障提供参考.     

荷泽电厂1号汽轮机组轴承振动诊断及处理

介绍了菏泽发电厂1号汽轮机组2号轴承振动异常的诊断过程及处理方法。从2002年10月2号轴承振动异常到2006年5月对转子加平衡块，经过了3年多不断地分析和处理，从增加轴承座动刚度着手，将2号轴振控制在报警值以下，基本上解决了2号轴承振动故障。     

330 MW汽轮机组高压转子突发性振动诊断及处理

6号汽轮发电组高压转子1号、2号轴承在运行中出现突发性轴振动大现象,振动趋势与负荷有关。查看汽轮机组在线振动分析系统,发现轴振动波动主要为25 Hz的半频分量,分析原因为蒸汽激振引起的轴承稳定性不足。利用停机机会解体检查轴承,发现轴承松动、轴承顶部间隙超标、轴承载荷异常等问题,通过采取重新校正等措施提高了轴承稳定性,机组启动后高压转子1号、2号轴承轴振降至合格范围,消除了6号机组高压转子汽流激振引发的突发性轴振大的问题。     

300MW机组汽动给水泵异常振动原因分析与处理

针对某300MW机组汽动给水泵轴承瓦振超标的问题,对汽动给水泵3号、4号轴承的瓦振现象进行了分析。给出了轴承瓦振的振动特征,分析了轴承瓦振超标的原因,由于支撑刚度弱和转子存在不平衡质量,使轴承产生了异常振动。采取了增加支撑和重新进行动平衡等措施,消除了轴承的异常振动现象。     

50MW供热机组汽流激振问题的分析及处理

某台50MW供热发电机组升至满负荷过程中振动大超过跳机值,根据振动现象、振动特征以及阀序调整试验,诊断为汽流激振。停机检查结果表明汽缸发生偏斜以及轴承稳定性降低是导致汽轮机发生汽流激振的主要原因。通过检修处理后,彻底解决了汽流激振问题,为类似振动故障处理提供了参考。     

大型发电机组轴承顶销松动故障诊断

汽轮机组轴承座松动,必然会引起轴承刚度的变化,而刚度的变化又会引起轴承座轴振和瓦振的变化.本文根据发电机组轴承刚度变化对机组振动的影响特性,成功诊断了一起300MW汽轮发电机组轴承座顶销松动故障.     

600MW汽轮发电机轴瓦振动分析

某厂600 MW火电汽轮发电机存在瓦振大、轴振小的问题。为解决此问题,应用振动力学理论建立大型旋转机械轴承-转子-支撑系统简化模型,结合该机组启动振动数据进行分析研究。结果表明,在柔性支撑系统下,轴振、瓦振关系受支撑动力特性影响较大,二者大小关系不再是简单的比例关系;轴、瓦振动均反映机组的振动性能,都应引起足够的重视,单独考虑其中一个都不合理;在升速过程中,机组轴振、瓦振大小变化比较复杂,其幅值大小及相位随转速增加而变化。最终得出,支撑刚度较弱是引起该机组出现瓦振大、轴振小现象的主要原因。     

基于深度傅里叶变换的调相机启停过程振动问题分析治理

振动问题是同步调相机组常见的问题之一.在某换流站2号调相机冷拖及惰转振动监测过程中,励端支撑系统水平方向的振动速度最高达到12.53 mm/s,超过跳机值11.8 mm/s.为确保后续整套启动及涉网试验安全,提出一种傅里叶变换和频响函数测试相结合的故障诊断方法.首先对旋转机械振动信号进行短时傅里叶变换,得到时频图,接着根据时频图结果开展频响函数测试分析.结果表明,在转子转速大于2700 r/min时,振动二倍频分量具有明显的信号特征,该分量是由励端支撑系统受迫振动频率与固有频率接近引起的.最终,通过将应急油箱支撑与轴承座进行改造分离,改变原有励磁支撑系统固有频率,能够在一定程度上降低冷拖及惰转过程中励端轴承座水平方向振动值.     

600 MW机组调试期间振动故障分析及处理

某台600 MW机组在调试期间发生高中压转子轴振不稳定、低压转子振动超标、发电机转子轴振超标和不稳定等故障.频谱分析结果表明,振动主要由转子不平衡、转子加工偏差、动静碰摩等所致.经实施高速动平衡、补偿、修刮轴承油挡等措施后,在满负荷工况下,除发电机转子7号轴承X方向,8号轴承X、Y方向外,其余轴承轴振均＜50 μm,瓦振＜15 μm,振动故障得以消除.     

汽轮发电机组开机调试过程中出现的不规则振动分析

某台350 MW汽轮发电机组开机调试过程中出现了不规则振动:①在发电机护环动平衡过程中,影响系数幅值和相位分散度很大:②机组升速过程中,发电机两侧轴振出现大量0～10 Hz的不规则低频分量.结合机组设备检修及动平衡过程对上述不规则现象进行了分析,指出压条松动和过大振动导致的轴瓦鹰损是造成不规则振动的原因.     

十里泉电厂125MW机组异常振动的诊断及处理

十里泉电厂2号机组大修后启动升速至3000r/min时,3号、5号轴承垂直振动分别为64μm、75μm,3X轴振为154μm,在处理过程中4X轴振变为182μm,低压转子和发电机转子的轴承振动相互影响,通过在发电机转子高速动平衡使振动水平达优良,本文详细介绍了该异常振动现象特征、原因及处理过程。     

引进型300MW机组易发振动故障的诊断处理方法

本文总结了上海引进型300MW汽轮发电机组轴系的组成、各个轴承的型式以及此种汽机目前在山东省的投运状况。重点介绍了该型机组的易发异常振动的8种故障,如低压转子不平衡、发电机励磁电机三支承转子不平衡、发电机转子热不平衡、低发对轮螺栓松动、低发对轮连接螺栓护板撕裂脱落、发电机励磁电机对轮螺栓紧力不足、励磁机轴晃动大、发电机密封瓦碰磨、副励磁机动静碰磨以及7号可倾轴瓦柱销磨损碎裂等。在分析总结调试及大小修该型机组异常振动的故障过程中,给出了各个故障形成原因、机理、表现、以及各个故障的最佳处理方法。最后提出该型机组低压转子4号轴承异常振动、发电机转子的热态不平衡、发电机励磁电机三支承平衡、励磁机7号轴承振动的处理方法,以及在机组大修及运行中应注意的问题。     

短路条件下变压器器身轴向振动的分析

基于电力变压器绕组轴向振动的质量弹簧系统数学模型并考虑变压器绕组的轴向振动的非线性特点,简化计算模型。本文针对一台型号为DFP1-240MVA/500kV的电力变压器,建立了由铁心、绕组、垫块、夹件和拉板组成的电力变压器器身振动的仿真模型,分别计算了变压器绕组和器身在不同轴向预紧力作用下的固有频率和振动型态,得出了轴向预紧力的变化与其轴向振动固有频率变化之间的关系。同时对短路条件下绕组轴向动态短路力所产生的器身轴向振动进行了分析。分析结果表明:短路时存在以50Hz和100Hz为基频的短路力,如果变压器的固有频率接近这两种频率及其倍频,会发生谐振,使振动位移变大,降低变压器结构的稳定性。     

某F级燃气机组振动故障诊断及处理

某F级燃机在热态停机惰走过程中1号轴承瓦振、2号轴承及3号轴承轴振均超过跳机保护值,根据振动现象、燃机转子结构及采集的振动数据,判断振动的主要原因为热态停机后的惰走过程中转子存在弯曲,通过2次现场动平衡校正,降低了机组热态停机后惰走时的振动异常.     

大型汽轮发电机组轴系振型识别方法研究

目前人们可以方便地获取汽轮发电机同承振动信号,但是对于轴系内部各点振动情况却知之甚少。本文运用模态分析理论建立了一转子－轴承系统轴系振型识别模型。该模型考虑了轴承等部件存在的各项异性、不同振动方向之间的耦合以及轴承座刚性等因素的影响,识别过程无需迭代求解。本文以转子试验台、连续转子和一大型汽轮发电机 为例详细分析了不同转速和不平衡分布下的识别结果。论文最后对本方法的误差进行了分析,以期提高本方法的实用性。     

Development of elements of the condition monitoring system of turbo generators of thermal power stations and nuclear power plants

The rationale is given for the improvement of the regulatory framework for the use of shaft sensors for the in-service condition monitoring of turbo generators and the development of control systems of shaft surfacing and misalignments of supports. A modern concept and a set of methods are proposed for the condition monitoring of the “shaft line–thrust bearing oil film–turbo generator supports” system elements based on the domestic COMPACS® technology. The system raw data are design, technology, installation, and operating parameters of the turbo generator as well as measured parameters of the absolute vibration of supports and mechanical quantities, relative displacements and relative vibration of the rotor teeth in accordance with GOST R 55263-2012. The precalculated shaft line assembly line in the cold state, the nominal parameters of rotor teeth positions on the dynamic equilibrium curve, the static and dynamic characteristics of the oil film of thrust bearings, and the shaft line stiffness matrix of unit support displacements have been introduced into the system. Using the COMPACS-T system, it is planned to measure positions and oscillations of rotor teeth, to count corresponding static and dynamic characteristics of the oil film, and the static and dynamic loads in the supports in real time. Using the obtained data, the system must determine the misalignments of supports and corrective alignments of rotors of coupling halves, voltages in rotor teeth, welds, and bolts of the coupling halves, and provide automatic conclusion if condition monitoring parameters correspond to standard values. A part of the methodological support for the proposed system is presented, including methods for determining static reactions of supports under load, the method for determining shaft line stiffness matrices, and the method for solving the inverse problem, i.e., the determination of the misalignments of the supports by measurements of rotor teeth relative positions in bearing housings. The procedure for calculating misalignments of turbo generator shaft line supports is set out.     

600MW超临界汽轮发电机组振动异常的原因分析及处理措施

针对600 MW超临界汽轮发电机组5号、7号、8号轴瓦及发电机定子振动异常的现象,通过振动数据的测试分析,认为低压转子不平衡、轴瓦顶隙超标是机组异常振动的原因,提出调整Ⅱ号低压转子动平衡和轴瓦顶部间隙、加固轴瓦下侧垫铁等处理措施.     

Results from Investigations of Torsional Vibration in Turbine Set Shaft Systems

The article generalizes the results obtained from investigations of torsional vibration in the shaft system of the T-175/210-12.8 turbine set installed at the Omsk CHPP-5 combined heat and power plant. Three different experimental methods were used to determine the lowest natural frequencies of torsional vibration excited in the shaft system when the barring gear is switched into operation, when the generator is synchronized with the grid, and in response to unsteady disturbances caused by the grid and by the turbine control and steam admission system. It is pointed out that the experimental values of the lowest natural frequencies (to the fourth one inclusively) determined using three different methods were found to be almost completely identical with one another, even though the shaft system was stopped in the experiments carried out according to one method and the shaft system rotated at the nominal speed in those carried out according to two other methods. The need to further develop the experimental methods for determining the highest natural frequencies is substantiated. The values of decrements for the first, third, and fourth natural torsional vibration modes are obtained. A conclusion is drawn from a comparison between the calculated and experimental data on the shaft system’s static twisting about the need to improve the mathematical models for calculating torsional vibration. The measurement procedure is described, and the specific features pertinent to the way in which torsional vibration manifests itself as a function of time and turbine set operating mode under the conditions of its long-term operation are considered. The fundamental measurement errors are analyzed, and their influence on the validity of measured parameters is evaluated. With an insignificant level of free and forced torsional vibrations set up under the normal conditions of turbine set and grid operation, it becomes possible to exclude this phenomenon from the list of main factors influencing the crack formation processes in low-pressure rotors. The importance of experimentally confirming the fact that the shaft system has been detuned from resonances at the 50 and 100 Hz excitation frequencies is pointed out.