外文翻译---液压机械装置振动问题处理.docx

原文solving vibration problems in hydraulic machineryabstract in the current paper, various cases of vibration problems detected in hydraulic machinery are presented.these cases were found during several years of vibration monitoring.the problems have been classified depending on their origin.for all of them,a systematic approach is given indicating the symptoms,the exciatation provoking them,the possibilities of amplification due to resonance and the remedies that have been applied.introductionvibration problems are common in hydraulic machinery.solving them helps to increase the machine life and to reduce maintenance costs.many cases have been found and solved during last years of monitoring (egusquiza 1998 and egusquiza et.2000).some of them have been due to design or mounting problems and others due to damage in machine elements.the cases presented here correspond to large machine with vertical shaft and rigid coupling.the problems found have been classified in the following types:type 1: excessive excitations of hydraulic origintype 2: hrust bearing problemstype 3: unbalance and misalignmenttype 4: electromagnetic problema methodology to solve these types of vibration problems is proposed in the paper.the steps to follow are indicated in fig,1.once the abnormal high vibration amplitudes have been detected through scheduled monitoring or machine malfunctioning.vibration analysis has to be proformde to identify the origin of the exciation provoking them.various techniques are available depending on the type of excitation under consideration,whether it is hydraulic,mechanic oreletromagnetion.the resulting high vibration levels may also be due to some type of resonace in the hydraulic system or in the mechanical compoents.therefore, this possibility must be checed before a correct diagnostic can be made.finally, remedies to the probiem are proposed and their success is confirmed comparing the vibrations after the modification with theprevious ones.type 1: excessive hydraulic excitation.typical vibrations of hydraulic origin are generated by rotor-stator interaction (rsi) and by cacitation.rsi is due to the interference between the runner blades and guide vanes.it can generate pressure pulsations of high amplitude in hydraulic machines.as a result,cracks in runners and excessive vibrationlevels in machine and piping can be produced.rsi identification is normally easy with spectral analysis of vibration.the problem is to know if the vibration is high due to the excitation itself,to a high hydraulic system is high due to the excitation resonse or to a runner/motor resonance.the vibration will be at a frequency fb given by: fb=n*f*zb=with maximum amplitude at the lowest diametrical mode excited k according to the following equation(tanaka 1990): n*zv+-k=m*zbanother typical phenomenon that can generate vibrations is cavitation.cavitation can take several forms and can result in vibrations,losses of performance and erosion. the most common type of cavitation is part load suige. its negative effects are usually mitigated by means of air entrapment in draft tube. on the other hand, inlet cacitation provoking erosion of runner blades is also a concern due to its destructive effects.here there are some examples for both types of hydraulic excitations.high vibrations in thecasingof a multistage pumpa multistage pump had high vibration levels in the casing during operation. the generated vibrations produced the burst of pipes and other elements. initially, it was thought that the vibrations were due to wear or damage(malfunctioning). as a result, the pump was completely dismantled and repaird. surprisingly, the vibration did not disappear.a general scheme of the pump is given in fig,2.during pumping the machine delivers a flow rate of 2.8 m3/s to a head of 935 m.other machine characteristics are listed in table 1.to understand the cause of the high vibertion levels some experimental measurements and a theoretical analysis were carried out.from the spectral analysis it was found that the vibration amplitude occurred at 150 hz which is fb(see fig,5).therefore,the vibration has a fluid dynamic origin and its high amplitude might be due to several reasons such as high rsi excitation due to design,mounting,damage,or resonance. in this case,the rsi analysis gives a pressure pulsation around the impeller with a diametrical mode k=-2 rotating in the opposite direction of the impeller st fb.first of all, the hydraulic system response of the return channel was calculated using a transfer matrix method. no frequencies around 150hz were found. furthermore, the analysis of the measurements(phase and amplitude variation)at different roating speeds neither showed resonance around the excitation frequency.modal analysis from the impacts done with an insteumented hammer was carried out to check the possibility of mechuanical resonance in the casing or impeller.for the casing,no resonance was found at around the frequency of 150 hz as it can be seen in fig,3. atheoretical analysis with fem gave similar results.the frequency response functions obtained from impacts in the impeller are indicated in fig,4. they show the presence of a mode at 472 hz. this mode was susceptible of excitation because it corresponded to diametrical mode 2. estimating a reduction factor of 0.45 to 0.5in order to consider the added mass and the casing boundary,the actual natural frequency would lie between 212 and 236 hz. so, there was a low probability of resonance. in fact, no high vibration levels were present at the bearings.finally, analyzing the phasing of the pressure pulsation and vibration in the casing, the diagnosis was that the high vibrations were caused by the interaction of the pressure pulsation inside the machine.in this case the solution was to change the relative position between impellers. after that, vibration was reduced considerably as it is shown in fig,5 where rsi vibrations before and after repair are compare.vibration on runner of a pump-turbine.in this case, the machine was a single stage reversible pump-turbine. its main characteristics are listed in table 2. this pump suffered form cracks in the impeller blades.for the zb and zv combinations, the diametrical mode k=-2 occurs at 2*fb(=140 hz) and should be the most important excitation. this is exactly what can be seen in the vibration spectrum shown in fig, 6.after analyzing the system response, checking the natural frequencies of runner and rotor, it was found that a rotor naturalfrequency was was close to the excitation frequency. as the system could not be changed physically, the solution of the problem was devised as modifying the runner lacation where cracks appeared to reduce stresses on blades.partial load surge and inlet cavitation erosion on blades of a francis turbine.this unie was a vertical shaft francis turbine operating up to a maximum output power of 65 mw. the total nominal flow rate and the net head were 57.5m3/s and 122.5m. the rest of characteristics are listed in table 3.in this case two problems had been detected. the first one was excessive vibration levels in the draft tube. the second one was advanced erosion on the suction side of the blades.in fig, 7 overall vibration levels measured for the entire range of power outputs indicate that amplitudes are more important during operation at partial loads. in the fig,8, spectral analysis of shaft displacement using proximity probes can be observed at 20 and 55mw. when operating at 20mw, a frequency peak at 0.27*ff predominates but disappears at 55 mw. this analysis indicated the presence of a hub rope in the draft tube at partial loads.spectral analysis is enough to detect partial load surge but it is not useful for other types of cavitation. this is the case of erosive cavitation is especially destructive.another technique to be used is to demodulate high frequency vibrations. in fig,9, the spectra of the envelope in the frequency band form 30k to 40k hz are plotted. again, at 20 mw the presence of the part load surge is well detected in the top of the fig.so,the possible remedies such as optimization of air injection fins are currently being analyzed. meanwhile, it has been recommended to avoid operation at loads below 35 mw.for the detection of erosive cavitation high frequency vibrations also had to be measured (escaler et al.2002). amplitude demodulation of high frequency bands, shown in the bottom of fig,9, indicated the presence of a pulsating cavity in the runner at fv at 55mw. the maximum amplitude of this peak was found at 60 mw, thus indicating the maximum cavitation aggressiveness. therefore, the solution consisted in limiting the time of operation around 60mw whenever possible. a refurbishing of the runner (new hydrodynamic design) is expected to be the solution in this case.type 2: thrust bearing problemsanother type of problem sometimes found in vertical shaft machines is rubbing in thrust bearing which can provoke its rapid destruction. in guide bearings,radial loading is usually low and they are not so affected.machines are prone to have friction during start-up and coast-down if only hydrodynamic lubrication exists. during operating, vibration can be generated especially when load on the pads is not evenly distributed. a possibility for detection is to install a vertical proximity probe or an absolute vibration sensor located next to the bearing pads in axial direction. the use of joint-time frequency analysis is very adequate to idenfuty such frictions that can occur in very short periods of time.at steady operation, spectral analysis helps to identify potential problems by looking at the pad passing frequency. for instance, in fig,10,this frequency disappears from the vibration signature after repairing the bearing. although detection is easy,to quantify the level of the damage is diffcult from the vibration signature and the peak amplitude. a proximity probe is convenient as well as oil analysis to complete the diagnostic.in fig, 11, another example is shown. the predominant peak at passing frequency indicates a damage in the bearing. its amplitude decreased significantly after repair.type 3: unbalance and misalignmentunbalance and especially misalignment are common problem in vertical shaft machine with rigid coupling.unbalbance detection and solution is not diffcult except when hydraulic or magnetic forces or a resonance zre involved in it. it is important to identify the type of unbalance before doing the repair. vibration measurements at different loads and with the machine idle are necessary. in fig, 12 the spectrum of a machine with unbalance produced by a blockage in runner channels can be observed. in this case, it was observed that the ff. another situation which is potentially dangerous is when a small part of the runner breaks off due to fatigue. here, the change in unbalance is mot large and damage is diffcult to detect.other cases difficilt to solve are when there is resonance with a rotor natural frequency. in fig,13，a case with misalignment can be observed. the first rotor lateral frequency is almost coincident with a times the rotating frequency. here the machine has some degree of misalignment which is enhanced by the resonance. in the top of the fig,14,the 2*ff peak has an rms amplitude around 1.2 mm/s when the machine operates at full load. meanwhile,in the bottom of the same fig,the same peak shows an amplitude of about 0.4 mm/s at 40% of the load.as the natural frequency varies depending on the machine load and other paraments, the vibration amplitudes change continuosly what makes diffcult to have an accurate ternd analysis in the monitoring. in this case the solution is not straightforward.rotordynamic analysis based in fem can be used to model the motor,to identify the type of resonance and to find a remedy. inaccuracies arise when simulating an installed machine due to the lack of exact geometrical data, bearing stiffness,and so on. therefore simulation must be checked with some experimental data. thisis rather complex because the natural frequency are diffcult to excite in a large machine. joint time frequency analysis can be used during transients or after impacting tha machine when in operation in order to identify them,as shown in fig,14.type 4: electromagnetic problemsthis type of problems are basically due to eccentricity or damage in generator.in fig,15,spectra of a machine before(front) and after excessive vibration in the generatoe are shown. the predominant vibration occurs at two times the electrical line frequency,in our case 100hz. after repair of the stator, where some damage was found,the vibration amplitude at 100hz was considerably reduced.翻译部分液压机械装置振动问题处理摘要：在本篇论文中，我们将讨论液压机械装置不同情况下的各种振动问题。这些情况在多年的正动监测中发现的。由于引起振动的本质不同我们将这些情况做下归类，对所有这些情况来说，用系统途径观察这些情况，超负荷运行引起这些振动。问题扩大的原因是由于机器的共振，一些措施已经应用于补救当中。简介：在液压机械装置中，振动问题非常普遍。解决振动问题可以延长机器寿命并且降低维修费用。通过去年的监测观察，许多振动问题被发现，并且解决了其中一部分问题。在发现的问题中一些是由于设计存在缺陷，一些是由于机器本身存在微小振动并且逐年增加的结果。还有一些是由于机器内部存在着某些零件的损坏登记。这所提到的例子大部分是有垂直通道和刚性连接的大型机器。我们可以将碰到的问题归为以下几类：类型1：液压负荷过大类型2：超载问题类型3：装置不平衡性和联接性类型4：电磁性问题论文中建议用统计分析的方法来解决这些类型的振动问题。图示为统计分析法的具体操作步骤。当通过定时监测器观察到不正常的大幅度振动或机器的不正常运转。振动分析将立即启动对所出现的问题的本质惊醒分析。许多的技术都可以运用到这些问题的研究中，不管机器是液压式还是动力机械式，或是电磁驱动的。导致大幅度振动的问题可能是液压系统经常出现的一些共振现象，或是一些机器零件的共振问题，因此这些的可能都必须再做出正确诊断前查清楚。最后，问题的补救措施已提供出来，下一步骤是将振动与修正前的结果进行比较进而得出正确的补救措施。类型1：液压负荷过大典型的液压振动的产生是音问机器旋转部分的相互影响，rsi是由发动机的叶片与引风道的相互影响产生。它可以在液压装置中产生大幅度的振动。rsi通常对振动进行光谱分析，问题在于知道振动是否是由于机器本身超负荷运行产生，对一个大型液压装置振动是否是由于动力装置或是旋转装置的共振引起。振动的频率由下式提供：fb=n*f*zb 。最大的幅度发生在最小的管道直径时。系数k由下列公式得到：n*zv+_k=m*zb。另一个可以产生振动的典型现象是气蚀。气蚀可以有多种形式出现。它有可能导致振动的发生。机器运转不正常就有可能是因为发生了腐蚀。最通常发生的腐蚀发生在部分负荷的高峰期。它的不利影响通常是由通风管道中空气流动引起的。另一面由于风道进口通风对动力装置叶片的腐蚀也是个不利的因素。下面介绍几个液压装置的例子供参考：多级抽水泵的高幅度振动多级抽水泵在工作时具有很大幅度的振动，由此产生的大幅度振动可能导致管道和其他的零件的破坏。首先，振动被认为是产生破坏或导致机器不能正常运转的问题。结果把水泵全部拆开进行修理，可是奇怪的是振动现象并不能消失。图2给出了多级水泵的大体结构。在抽水的过程中，水泵以2.8m/s的流速向935米高的地方抽水，水泵的其他性能参数在表格1中列出。为了搞清楚多级水泵大幅度振动的问题，进行了一些实验测量和理论分析。从光谱分析中可以知道，振动幅度增大发生在振动频率为150hz时，因此其振动的本质是由一些不稳定的流量产生的。多级水泵的高幅度振动可以归结为以下几个原因，例如是设计时的高rsi，逐渐增加的振动频率，零部件的损坏，共振等。在此例中，rsi分析在驱动器旁边测量了其振动压力，它是在外径模型系数k=-2时，频率为fb并且相反旋转方向测得的。首先，液压系统回水管道的反应是使用矩阵转换的方法测得的，在150赫兹左右其他频率都不可能测到。其次，对在不同转动速度下的测量数据进行分析也没有在过高负荷的频率附近产生共振。从电击锤碰撞的分析中确定机械共振的可能性。在这次实验分析中，频率在150赫兹附近没有发现共振现象，我们可以从图3中看出来。用fem理论分析可以得到相同的结论。图4说明了在驱动装置中通过撞击得到的频率反应函数。两