李旋翻译输油管道中往复式三柱塞泵的震动与失效问题分析

输油管道中往复式三柱塞泵的振动与失效问题分析J.C.W ACHEL, 总裁ASME( American Society of Mechanical Engineers美国机械工程师协会)成员工程动力公司F.R.SZENASI, 资深项目工程师ASME( American Society of Mechanical Engineers美国机械工程师协会)成员圣 .安东尼奥 ,德克萨斯S.C.DENISON, 产品工程顾问AIME( 美国采矿、冶金及石油工程师学会 )/SPE(Society of Petroleum Engineer 石油工程师学会 ) 成员太耐克石油勘探生产公司休斯顿德克萨斯ABSTRACTAn analysis was made to identify the causes of vibration and failure problems with the piping and reciprocating pump internals on an oil pipeline pump station. A field investigation was made to obtain vibrations and pulsations over the entire range of plant operating conditions. The data showed that cavitation was present at nearly all operating conditions due to the high pulsations in the suction system. The discharge system experienced high vibrations and piping failures due to the ineffective of the accumulator. An acoustical analysis of the suction and discharge system was made to design the optimum acoustical filter systems to alleviate the problems. The acoustical analyses were performed with a digital computer program which predicts the acoustical resonant frequencies and the pulsation amplitudes over the speed range. This paper discusses the investigations and gives recommendations for prevention of these types of problems in the future.摘要本文完成了对某泵站输油管线振动原因及往复泵内部部件失效原因的分析。为获得全部运行工况下的振动量和脉动值, 本文进行了现场研究。现场数据表明：由于吸入系统中存在较高的液流脉动 ,在近乎所有的运行工况下中存在气穴现象。由于排出系统的缓冲器无效，故排油管线存在剧烈振动和管路失效。为了设计出优化的声学滤波系统来缓解上述问题, 本文对吸、排系统进行了声学分析。声学分析由数字计算机程序完成，该程序预测出在全部速度范围内的声学共振频率和脉动幅值。本文还讨论了这些研究结果，并给出了若干建议以防止诸如此类问题的发生。INTRODUCTIONProblems were experienced with four triplex reciprocating crude oil pumps operating in parallel at the Dina Pumping Station located in Colombia (Figure l). The pumps had a rated speed of 275 rpm with a capacity of 388 gallons per minute. The nominal suction pressure was 60 psig (414 kPa) and the discharge pressure was 1800 psig (12400 kPa). The Delrin pump valves had repeated fatigue failures beginning three months after startup. The discharge valve disks were replaced with steel and the Delrin disks used on the suction valves were replaced every 90 days to avoid fatigue failures. Valve failures were controlled after the first 9 months of station operation. For the firstfour months there were no pull rod failures; however, there have been 18 failures in the followingyear and a half. Many of these failures required replacement of the crosshead, the guide ways and on two occasions a broken or bent connecting rod. The suction and discharge piping systems vibrated excessively, resulting in several piping fatigue failures. Attempts to control the piping vibrations with pipe clamps and additional supports were unsuccessful.引言问题发生在位于哥伦比亚的迪那泵站，四台三柱塞式往复原油泵并机运行（图1），往复泵的额定转速为 275rpm，额定排量为每分钟 388 加仑，正常的吸入压力为 414 千帕，排出压力为 12400 千帕。聚脂阀片在启动之后三个月内反复出现疲劳失效。排放阀片由钢质替代， 用于吸入阀片每隔九十天进行更换以防疲劳失效。在油泵站运行的前九个月， 阀的失效得以控制， 因为前四个月不会有拉杆失效问题，然而，在接下来的一年半中出现了十八次这样的失效。 许多这样的拉杆失效需更换十字头，导轨，有两种情况出现了连接杆断裂或弯曲。吸、排输油管线振动过大，导致若干输油管线疲劳失效。 试图用管卡和附加支撑来控制输油管线振动往往无效。The four pumps had a common suction header supplied by a charge pump which was capable of supplying pressures up 0 90 psi (621 kPa) .The discharge of the four pumps fed into a common header which connected to the main pipeline. The original piping design included bladder-typeaccumulators on both the suction and discharge.四台泵共用一只吸入集油管，用一台排油泵为该集油管提供 621 千帕吸入压力。四台泵共用一只排出集油管， 排出集油管与主管道相连， 原设计中吸入、 排出管路均包括有气囊式缓冲器。It was difficultto keep the pumps running smoothly since constant maintenance was needed tokeep the accumulator or bladder pressures charged to approximately60 t0 70 percent of linepressure. The static discharge pressure could change from 700 psig (4826kPa)to more than1600 psig (11032 kPa) in a few minutes if the down-line booster station went down. When this happened, the accumulator was ineffective.要维持这些油泵平稳运行十分困难， 因为缓冲器中的气囊需要持续充压以达到约百分之六十至七十的管路压力。 如果下游增压站运行工况发生变化， 排出静压在几分钟之内能从 4826 千帕增加到 11032 千帕以上。若这种情况发生，缓冲器便会失效。The cost of the parts and labor that could be attributed to this problem was in excess of $500,000. Tenneco, the piping designer and the pump manufacturer began a study to determine the cause or causes of the vibrations and failures. However, the complex relationship of the system variables made it difficult to develop definite conclusions.由于这些问题导致的另部件成本和人工费用多达50 万美元。在 太耐克公司内部（ Tenneco）,管路设计人员和制造商已经展开了对振动及管路失效原因的探究，然而，系统中变量之间的关系复杂使产生一个明确的结论很难。There were several changes during this phase in a vibrations and reduce the made in the piping system attempt improve the failure. These included changing the piping (at the recommendation of the accumulator vendor) so that the flow would be directed at the bladder. This piping modification did not improve the pulsation characteristics of the system.在此期间 ,致力于改善振动以及减少管路失效对于管道系统也做过许多整改工作。包括调整管道使液体能直接通向气囊（在缓冲器供应商的建议下） 。此次管道改造也未能改善系统的脉动特性。Another modification which was tried on the suction side of pumps I and 3 was the replacement of the bladder- type accumulators with nitrogen- charged, flow-through accumulators (Figure l). No noticeable improvements were observed after these changes were implemented.另一种致力于泵的吸入端 1 和 3 的修正是气囊式充氮缓冲器和流通式缓冲器的置换。但是在这些修正之后泵的脉动特性并没有很显见的改善。FIGURE l.Pump Piping Layout Showing PressureMeasurement Locations图 1. 泵管路布置的压力测量位置视图。The severity of the problems brought the basic design of the system into question since the suction and discharge lead lines from the headers to the pump manifold were shorter than normal for most pipeline stations. The pumps were located on 16 foot (4.88 m) centers with the suction and discharge headers located 10 to 12 feet (3.05 t0 3.66 m) away from the pump flanges.自从从源头到泵体的吸入和排除主管道在大部分管路位置比正常的变得更短之后，系统基本设计的问题严重性出现了。泵体位于比吸入和排出端高 16 英尺的位置，而吸入和排出端离泵的法兰大约 10 到 12 英尺的位置。The station capacity was 39900 barrel per day (264m3/h) when the pumps were at their rate capacity of 388 gallons per minute (88 m3/h) .this results in a fluid velocity of 3.3 ft/s (1 m/s) inthe 12 inch schedule 40 suction manifold and 6.9 ft/s (2.1 m/s) in the10 inch schedule XS discharge manifold. The flow velocities in the individual pump piping were 1.1 ft/s (0.34 m/s) inthe 12 inch standard weight suction pipe and 2.7 ft/s (0.82 ra/s) in the 8 inch extra heavy discharge pipe.当泵的功率容量达到 388 加仑每分钟（88 立方米每小时）时，水站的容量是 39900 桶每天（264 立方米每小时）。这个导致了流体的速度为 3.3 英尺每秒（一米每秒）在 12 英寸的吸入室，和 6.9 英尺每秒（ 2.1 米每秒）在 10 英寸的排出室。泵的独立管道流体速度是 1.1 英尺每秒（ 0.34 米每秒）在 12 英寸标准重量的吸入管道和 2.7 英尺每秒（ 0.82 米秒）在 8 英尺的附加重排出管道。Engineering Dynamics Incorporated (EDI) was requested to investigate and make recommendations to alleviate the problems.The first step in the analysis was to model the acoustical characteristics of the piping systems on a digital computer program to define the expected pulsation resonances and the overall amplitudes in the suction and discharge piping.A detailed field invesrigation was then made to evaluate the pulsation and vibration characteristics of the pumps. Solutions were then developed to eliminate the problems.Engineering dynamics incorporated 被要求去调查和对其中一种问题作推荐， 分析第一步就是在电子计算机程序上模仿其声音特性并在吸入和排出管道中定义其期望的脉动相应和最大振幅。 为了评估泵的脉动特性和振动特性， 做了一个很详细的广泛的调查。研究出消除这些问题的解决方案。FIELD INVESTIGATIONInstrumentation And Test ProceduresThe instrumenration and data acquisition system used to determine the pulsation and vibration characteristics are shown in Figure 2.Piezoelectric pressure transducers and accelerometers were used to measure che presaure pulsations and the vibrations. A sketch of the pump suction and discharge piping illustrating some of the pressure test points are shown in Figure l. The pulsation and vibration signals were analyzed for frequency conr:ent with a two channel Hewlett-Packard 3582A FFT analyzer and documented on a HP 7470A digital plorter. The analyzer and instruments were controlled by an Apple II+ microcomputer using software written specially for analyzing vibration and pulsation data. Torsional vibrations were measured with a HBM torsiograph mounted on the stub end of che pinion gear shaft on pump l.现场调查研究仪器和测试流程如图 2 所示仪器和数据获得系统是用来用来测量脉动和振动特性。 压力电压转换器和加速极用来测量压力， 脉动和振动。泵的吸入和排出结构的一些测量点如图1 所示。频率容量的脉动和振动信号的分析和测量是由双通道hewlett-packard3582A FFT 分析器和一台 HP 7470A 数字计量仪测量的。这个分析器和仪器是由一台苹果的 II+ 微机利用专门为分析振动和脉动数据的软件控制。扭转振动式由一台在泵 1 转轴齿轮末端转轴测量仪所测量的。2 Channel FFT Analyzer 通道 FFT 分析仪Micro-Computer 微型计算机Floppy Di3k Drive软盘 Di3k 车道Digital Plotter数字绘图仪Tuneable Filters 可调滤波器2 Channel Oscilloscope2 通道示波器Transducer S18nal Conditioner and Power Supply8 Channel FM Tape Recorder8 频道 FM 录音机Function Generator 函数发生器Strain Gage Amplifier and Frequency DemodulatorFIGURE 2.Data Acquisition System传感器 S18nal 调节器和能源供应应变放大器和频率解调器数据咨询系统Vibration and Pulsation TestingThe initial vibration surveys revealed high vibration amplitudes on the piping,indicating large excitation forces present in the piping system.analysis of the present pulsation waveforms revealed severe cavitation in the suction piping system.This cavitation was the source of the high energe causing the high piping vibration, valve failures, and pump part failures.震动与脉动测试最初的震动调查最初的调查显示， 在管道中的高震动频率， 预示着大的集电冲力存在于管道系统分析， 现有的震动波形揭示了吸入系统中严重的气穴现象， 高能量引起的高震动，阀门失效以及部分管道失效皆源于此种气穴现象。Cavitation气穴For liquid reciprocating pumps, the static pressure in the suction system must be adequte to compensate for acceleration head,and the pulsations present in the system.This ensures that the pressure of the oil was less than 2 psia.when pulsations exit in a system,they will be added to the static pressure and a negative peak which will be substracted from the stastic pressure.If thenegative peak of the pulsation, when subtracted from the static pressure, reaches the vapor pressure, the fluid will cavitate,resulting in high pressure spikes as the liquid vaporizes and then collapses as the pressure increases above the vapor pressure.对于液态往复式泵来说， 吸入系统中的静压力必须足够弥补系统中的加速头和脉动。这就保证了石油的压力小于 2psia.当脉动存在于系统中时，它们将被添加到静态压力和负峰之上， 负峰是从静压力中扣除的， 当脉动的负峰从静压力下扣除时达到饱和压力， 那么液体将会产生气穴， 然后在液体蒸发时产生峰值高压， 最后在蒸汽压力之上的压力便会崩溃。To illustrate the effects of cavitation, consider the pressure-time wave shown in Figure 3 which shows that cavitation occurs on the suction stroke.Note that when the cavitation portion of the waveform is expanded,the pressure spikes are approximately 0.00025 seconds.The pressure of cavitation can usually be observed on the complex wave since pulsations,which “squre-off ”at the trough of the waves when the vapor pressure is reached. The type of data to substantiate cavitation(Figure 4) illustrates the squaring-off of the wave, followed by the sharp spiking characteristic of severe cavitation. This data, taken on pump unit 1,showed pressure spikes of 600 psi(4137 kpa).为了解释气穴的影响，考虑到柱塞压力波时间会体现在图3，这就表明空化现象会发生在吸力冲程中，当波形中气穴比例扩大时，高压尖峰将会接近 0.00025 每秒，通常在源自脉动的复杂的电波中能观测到气穴。其中“平方压力”出现在当蒸汽压力到达波槽的情况下。显示气穴现象的数据类型解释了电波的平方关闭，随之而来的是严重的气穴现象下尖锐的摩擦。 这些取自管道单元 1 的数据显示了高达 600psi 的高压尖峰。FIGURE 3.Pump Plunger Pressure-time WaveShowing Cavitation泵柱塞压力与时间关系波形显示气穴现象FIGURE 4. Cavitation Caused By High Pulsations高脉动引起的气穴现象The effect of the statuc pressure on the cavitation was investigated by raising the suction pressure to the maximum possible (90 psig/621 kpa).The increase in suction pressure alone was not sufficient to eliminate the cavitation. Severe pulsations were found with levels in excess of 200 psi peak-to-peak. At a suction pressure of 76.5 psia , pulsations of approximately 75 psi zero-peak are required to cause cavitation. This value is obtained by substracting the negative pulsation peak from the static pressure. Since pulsations greater than 75 psi were always present at the higher speeds, cavitation always occurred.通过把吸入压力提高到最大可能值来调查气穴现象下静态压力的影响。 单独靠增加吸入压力不足以消除气穴现象。平均增加 200psi 对峰峰值时会有严重的脉动。在吸入压力为 76.5psi 和将近 75psi 零峰值情况下的脉动都会引起气穴现象。 通过从静压力减去脉动负峰从而得到这个值。由于高于 75psi 的脉动通常存在于更高的速度条件下，因此气穴现象时有发生。In the presence of caviLation, it is practically impossible to evaluate the influence of variables, such as the effect of other units, speeds and the accumulator design. Obviously, a reduction of the pressure pulsations was necessary in order to obtain meaningful test data on the units.气穴现象存在时，实际操作上不能去评估各种影响，例如其他单位，速度，以及蓄能器设计的影响。 显然，为了得到单元中有意义的测验数据， 减少高压震动很有必要。AcousticalResonances .声学共振The major succion pulsation -components were at frequencies near ll0 to 150 Hz with pulsation amplitudes of approximately 100 - 150 psi (689 - 1034 kPa) peak-to-peak, which, whencombined with the pulsation at the lower pump harmonics, caused the overall static pressure to drop below the vapor prassure. It l was detemined that acoustical resonances were causing the high amplitude pulsations. Acoustical resonances amplify the pulsations whenever one of theharmonics of the pump speed passes through the resonant frequency. The acoustical resonance at 130 Hz was a quarter-wave resonance of the suction pipe and was associated with the 9 foot (2.74 m) length from the end of the suction manifold to the accumulator.主要的吸入震动成分，频率在近 110-150Hz，震动振幅接近 100-150psi 对峰值，加上在较低的脉动泵和声学， 所造成的静态压力降至蒸汽压力， 毫无疑问，脉动振幅讲引起声学共振。 无论何时一个泵的和声学的速度超过共振频率， 声学共振都将扩大振动。 130Hz 的声共振是一种吸入管道季波的共振， 并与来自累加到蓄能器的九英尺长的吸管尾端联结。When an acoustical resonance is encountered in a system, the pressure pulsations can be reducedby eliminatingthe resonance orby attenuating the amplitudes throughthe addition of aresistiveelement, such as an orifice. Therefore, an orifice plate was instaLled at the suction flangein an attemptto attenuate the pulsation amplitudes and possibly movethe acoustical naturalfrequency. Adiameter ratio (orifice diameter to inside diameter ofpipe)of approximately0.4was used to ensure a significantacoustical effect. When the orificeplatewas installed,thepulsations were reduced; however,the reduct:ion was not sufficient to completely eliminate thecavitation.当系统中遭遇声学共振， 通过消除共振或是通过增加诸如孔之类的电阻元件来减少振幅便可以减少振幅。于是，在吸力凸缘上安装一个孔板试图减少脉动振幅，并可能移除声学固有频率。一个大约 0.4 的直径比（孔口管道内直径）被用来确保重要的声学效果。 当孔板被安装后， 脉动即会减少， 但是减少的程度还不足以完全消除气穴现象。Interaction With Other Pumps.与其他泵的相互作用All the other pumps were shut down and pump I was run to determine if the cavitation was caused by interaction with the other pumps or was a function of the individual piping design. These tests indicated that the pulsations were caused by the individual pumps and that the major factor was the acoustical resonances near 130 Hz. This test also gave evidence that the location of the pumpin the manifold system was not a major factor in the cavitation. This is verified by the factthatcavitation occurred on units I and 3 at the exact same speed under the same operating conditions. Units l and 3 are separated by 32 feet (9.75 m) with unit 2 midway between them. If the locationof the pump in the header was a prime facCor, t:here would have been different pulsation andcavitation characteris tics .To further investigate the interaction of the other pumps, tests weremade with pump I on the verge of cavitation and the adjacent unit 2 was swept through the entirespeed range to determine if it affected the speed at which cavitation occurred. This test showedthat the adjacent unit: did not influence the cavitation.Inan attemptto deterraine whether theacoustical resonance was associated with a piping length from the other units, the suction blockvalve was pinched momentarily to see if a pressure drop taken on the upstream side of the accumulator would affect the resonances in the 130 Hz range. The pressure drop of approximately l0 psi in the block valve did not have a significant effect.所有其他的泵都关闭后， 泵 1 继续运行，通过与其他泵相互作用或是个人管道设计的作用来决定气穴现象会否发生。 这些测试表明脉动是由单个泵引起的且其主要原因是声学共振接近 130 赫兹。此项测试也证明了在多方面系统中泵的位置不是气穴现象的主要原因。 同一操作条件下， 气穴现象以完全相同的速度发生在单元 1 和 3，通过这一事实，此项测试也得以证实。单元 1 和 3 被单元 2 以 32 英尺的距离在二者中间隔开。 如果头泵的位置是一个主要原因， 那么将会有不同的脉动和气穴特征。 为了进一步研究与其他泵的相互作用， 进行了相关测试， 在气穴边缘以及与其毗邻的单元 2 覆盖了整个速度范围来确定是否是它影响了气穴发生时的速度。 此项测试还表明与其毗邻的单元并没有影响到气穴发生。 为了摸清声学共振与其他管道长度是否有关系， 吸座阀被瞬间压紧来判断蓄能器的上流一侧压力骤降时会否影响到 130 赫兹范围内的共振。座阀上接近 10psi 的压力下降并无决定性的影响。Finalng Testing.最终测试After the orifice plate was installed and the nitrogen-charged accumulator bottle on the suction system had the maximum gas charge, the cavitation was elitainated over much of the speed range making it possible to study the effect of varying system parameters. The normalprocedure for the testing was to establish a set of steady-seate conditions, (such as suction pressure, gas volume in the bottle, or charge pressure in the bladder accumulator, peeds on the other pumps, etc.) and then change the pump speed from 190 rpm to 290 rpm. During the speed run, thepulsations in the sucCion and discharge piping were tape recorded for later evaluation. The resulting data presentation for the speed variation is given in Figure 5, showing the harmonics of pump speed pulsation pressures in the suction manifold of pump 3 over the speed range. The data shows that the primary cause of the cavitation was the high level pulsations at the acoustical natural frequencies in the system near 130 and 140 Hz which were excited by the 2lst through the 30th harmonics of pump speed.孔板被安装后， 且安装在吸入系统中的氮充电式蓄能器得到最优气体充电后， 气穴现象将在更大范围内消除， 这就使得研究多变系统指标的作用成为可能。 正常的测试程序是建立一系列的稳座条件 （如吸入压力， 瓶中气体容量， 气囊式蓄能器中的充电压力，其他泵的速度等等）然后把泵的速度从 190 转切换到 290 转。在此速度运行中， 吸排管道的震动被磁带录下用以稍后评估。 这些数据表明气穴的首要原因是在将近 130 到 140 赫兹的兴奋是由为 21 泵的额速度通过第三十次谐波系统的声学固有频率高的水平脉动。FIGURE5.SpeedRasterOfPumpSuctionPulsations泵吸脉动的速度光栅Speed Effects.速度因素The effect of speed on cavitation can be seen in Figure 6 which gives the complex pressure wave for speeds from 220 t0 270 rpm for a suction pressure of 60 psig (414 kPa). Pulsations generally increase with speed unless there are acoustical resonances. As shown, when the speed increased above 250 rpm, the pulsations increased to the point chat the negative pressure pulsation amplitude was the capor pressure and the wave became flattened on the trough. As the speed was further increacedthe cavitation became more severe .气穴现象中速度的影响可以在图 6 中看到，图 6 给出了速度从 220 到 270 的复杂的压力波，吸入压力为 60psi，脉动会速度增长除非有声学共振。如数据显示，当速度增加到高于 250 转时，脉动增加到一个点， 负压脉动振幅是蒸汽压力和波形在槽中变得平稳。由于速度变得更快，气穴现象便会变得更加严重。FIGURE 6.Complex Wave Of Pressure Pulsation VersusSpeed For Suction Pressure Of 60 Psig压力脉动对应的复杂波形60psig 吸入压力的速度Static Pressure Effects.静态压力的影响When the static suction pressure was inreased to 90 psig, the pulsation amplitudes were reduced and the unit could be run at 280 rpm without cavitation (Figure 7). The higher suction pressure seemed to inhibit the amplitude of the pulsations. The results of these tests indicated that_ the cavitarion could be minimized by increasing the suction pressure to the maximum possible, installing an orifice plate to reduce the pulsation amplitudes, and ensuring that the accumulator was properly charged.当静态吸入压力达到 90psi 时，脉动振幅便会减少，机组能运行至 280 转，且没有气穴（见图 7）.较高的吸气压力似乎抑制了脉动振幅。这些测试的结果表明，通过增加吸气压力到最大可能值， 安装一个孔板来减少脉动振幅， 并确保蓄能器供电充足，那么，气穴现象的发生就可以最小化。