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1、 輸油管道中往復式三柱塞泵的振動與失效問題分析 J.C.W ACHEL, 總裁 ASME( American Society of Mechanical Engineers 美國機械工程師協(xié)會 ) 成員 工程動力公司 F.R.SZENASI, 資深項目工程師 ASME( American Society of Mechanical Engineers 美國機械工程師協(xié)會 ) 成員 圣 .安東尼奧 ,德克薩斯 S.C.DENISON, 產(chǎn)品工程顧問 AIME( 美國采礦、冶金及石油工程師學會 )/SPE(Society of

2、Petroleum Engineer 石油工程師學會 ) 成員 太耐克石油勘探生產(chǎn)公司 休斯頓 德克薩斯 ABSTRACT An 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 pulsation

3、s 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 a

4、coustical 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 s

5、peed range. This paper discusses the investigations and gives recommendations for prevention of these types of problems in the future. 摘要 本文完成了對某泵站輸油管線振動原因及往復泵內(nèi)部部件失效原因的分析。 為獲得全部運行 工況下的振動量和脈動值 , 本文進行了現(xiàn)場研究?，F(xiàn)場數(shù)據(jù)表明：由于吸入系統(tǒng)中存在較高 的液流脈動 ,在近乎所有的運行工況下中存在氣穴現(xiàn)象。由于排出系統(tǒng)的緩沖器無效，故排 油管線存在劇烈振動和管路失效。 為了設

6、計出優(yōu)化的聲學濾波系統(tǒng)來緩解上述問題 , 本文對 吸、排系統(tǒng)進行了聲學分析。 聲學分析由數(shù)字計算機程序完成， 該程序預測出在全部速度范 圍內(nèi)的聲學共振頻率和脈動幅值。 本文還討論了這些研究結(jié)果， 并給出了若干建議以防止諸 如此類問題的發(fā)生。 INTRODUCTION Problems were experienced with four triplex reciprocating crude oil pumps operating in parallel at the Dina Pumping Station located in Colombia (Fi

7、gure 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

8、 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 first four months there were no pull rod failures; however, there have

9、 been 18 failures in the following year 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 failure

10、s. Attempts to control the piping vibrations with pipe clamps and additional supports were unsuccessful. 引言 問題發(fā)生在位于哥倫比亞的迪那泵站，四臺三柱塞式往復原油泵并機運行（圖 1），往復泵的額定轉(zhuǎn)速為 275rpm，額定排量為每分鐘 388 加侖，正常的吸入壓力為 414 千帕，排出壓力為 12400 千帕。聚脂閥片在啟動之后三個月內(nèi)反復出現(xiàn)疲勞失效。排放閥片由鋼質(zhì)替代， 用于吸入閥片每隔九十天進行更換以防疲勞失效。在油泵站運行的前九個月， 閥的失效得以控制， 因為前

11、四個月不會有拉桿失效問題，然而，在接下來的一年半中出現(xiàn)了十八次這樣的失效。 許多這樣的拉桿失效需更換十字頭，導軌，有兩種情況出現(xiàn)了連接桿斷裂或彎曲。吸、排輸油管線振動過大，導致若干輸油管線疲勞失效。 試圖用管卡和附加支撐來控制輸油管線振動往往無效。 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 int

12、o a common header which connected to the main pipeline. The original piping design included bladder-type accumulators on both the suction and discharge. 四臺泵共用一只吸入集油管，用一臺排油泵為該集油管提供 621 千帕吸入壓力。四臺泵共用一只排出集油管， 排出集油管與主管道相連， 原設計中吸入、 排出管路均包括有氣囊式緩沖器。 It was difficult to keep the pumps running smooth

13、ly since constant maintenance was needed to keep the accumulator or bladder pressures charged to approximately 60 t0 70 percent of line pressure. The static discharge pressure could change from 700 psig (4826 kPa) to more than 1600 psig (11032 kPa) in a few minutes if the down-line boost

14、er station went down. When this happened, the accumulator was ineffective. 要維持這些油泵平穩(wěn)運行十分困難， 因為緩沖器中的氣囊需要持續(xù)充壓以達到約百分之六十至七十的管路壓力。 如果下游增壓站運行工況發(fā)生變化， 排出靜壓在幾分鐘之內(nèi)能從 4826 千帕增加到 11032 千帕以上。若這種情況發(fā)生，緩沖器便會失效。 The cost of the parts and labor that could be attributed to this problem was in excess of $500,000.

15、 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. 由于這些問題導致的另部件成本和人工費用多達 50 萬美元。在 太耐克公司內(nèi)部 （ Tenneco）,

16、管路設計人員和制造商已經(jīng)展開了對振動及管路失效原因的探究，然而，系統(tǒng)中變量之間的關(guān)系復雜使產(chǎn)生一個明確的結(jié)論很難。 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

17、be directed at the bladder. This piping modification did not improve the pulsation characteristics of the system. 在此期間 ,致力于改善振動以及減少管路失效對于管道系統(tǒng)也做過許多整改工 作。包括調(diào)整管道使液體能直接通向氣囊（在緩沖器供應商的建議下） 。此次管道改造也未能改善系統(tǒng)的脈動特性。 Another modification which was tried on the suction side of pumps I and 3 was

18、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. 另一種致力于泵的吸入端 1 和 3 的修正是氣囊式充氮緩沖器和流通式緩沖器的置換。但是在這些修正之后泵的脈動特性并沒有很顯見的改善。

19、 FIGURE l. Pump Piping Layout Showing Pressure Measurement Locations 圖 1.. 泵管路布置的壓力測量位置視圖。 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 manifo

20、ld 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. 自從從源頭到泵體的吸入和排除主管道在大部分管路位置比正常的變得更短之 后，系統(tǒng)基本設計的問題嚴重性出現(xiàn)了。泵體位于比吸入和排出端高 16 英尺的位置，而吸入和排出

21、端離泵的法蘭大約 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) in the 12 inch schedule 40 suction manifold and 6.9 ft/s (2.1 m/s) in the10 inch

22、schedule XS discharge manifold. The flow velocities in the individual pump piping were 1.1 ft/s (0.34 m/s) in the 12 inch standard weight suction pipe and 2.7 ft/s (0.82 ra/s) in the 8 inch extra heavy discharge pipe. 當泵的功率容量達到 388 加侖每分鐘（88 立方米每小時）時，水站的容量是 39900 桶每天（264 立方米每小時）。這個導致了流體的速度為 3.3

23、 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。泵的獨立管道流體速度是 1.1 英尺每秒（ 0.34 米每秒）在 12 英寸標準重量的吸入管道和 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 analy

24、sis 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 characte

25、ristics of the pumps. Solutions were then developed to eliminate the problems. Engineering dynamics incorporated 被要求去調(diào)查和對其中一種問題作推薦， 分析第一步就是在電子計算機程序上模仿其聲音特性并在吸入和排出管道中定義其 期望的脈動相應和最大振幅。 為了評估泵的脈動特性和振動特性， 做了一個很詳細的廣泛的調(diào)查。研究出消除這些問題的解決方案。 FIELD INVESTIGATION Instrumentation And Test Procedure

26、s The 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 p

27、iping 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 A

28、pple 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. 現(xiàn)場調(diào)查研究 儀器和測試流程 如圖 2 所示儀器和數(shù)據(jù)獲得系統(tǒng)是用來用來測量脈動和振動特性。 壓力電壓轉(zhuǎn)換器和加速極用來測量壓力， 脈動和振動。泵

29、的吸入和排出結(jié)構(gòu)的一些測量點如圖 1 所示。頻率容量的脈動和振動信號的分析和測量是由雙通道 hewlett-packard 3582A FFT 分析器和一臺 HP 7470A 數(shù)字計量儀測量的。這個分析器和儀器是由 一臺蘋果的 II+ 微機利用專門為分析振動和脈動數(shù)據(jù)的軟件控制。 扭轉(zhuǎn)振動式由 一臺在泵 1 轉(zhuǎn)軸齒輪末端轉(zhuǎn)軸測量儀所測量的。 2 Channel FFT Analyzer 通道 FFT 分析儀

30、 Micro-Computer 微型計算機 Floppy Di3k Drive 軟盤 Di3k 車道 Digital Plotter 數(shù)字繪圖儀 Tuneable Filters 可調(diào)濾波器 2 Channel Oscilloscope2 通道示波器 Transducer S18nal Conditioner and Power Supply 8 Channel FM Tape Recorder8 頻道 FM 錄音機 Function Generator 函數(shù)發(fā)生器 Strain Gage Amplifier and Frequency

31、Demodulator FIGURE 2. Data Acquisition System  傳感器 S18nal 調(diào)節(jié)器和能源供應 應變放大器和頻率解調(diào)器 數(shù)據(jù)咨詢系統(tǒng) Vibration and Pulsation Testing The initial vibration surveys revealed h

32、igh 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 failur

33、es, and pump part failures. 震動與脈動測試 最初的震動調(diào)查最初的調(diào)查顯示， 在管道中的高震動頻率， 預示著大的集電沖力存在于管道系統(tǒng)分析， 現(xiàn)有的震動波形揭示了吸入系統(tǒng)中嚴重的氣穴現(xiàn)象， 高能量引起的高震動，閥門失效以及部分管道失效皆源于此種氣穴現(xiàn)象。 Cavitation 氣穴 For liquid reciprocating pumps, the static pressure in the suction system must be adequte to compensate for acceleration hea

34、d,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 the negative peak of the pulsatio

35、n, 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. 對于液態(tài)往復式泵來說， 吸入系統(tǒng)中的靜壓力必須足夠彌補系統(tǒng)中的加速頭和脈動。這就保證了石油的壓力小于 2psia.當脈動存在于系統(tǒng)中時，它們將

36、被添加到靜態(tài)壓力和負峰之上， 負峰是從靜壓力中扣除的， 當脈動的負峰從靜壓力下扣除時達到飽和壓力， 那么液體將會產(chǎn)生氣穴， 然后在液體蒸發(fā)時產(chǎn)生峰值高壓， 最后在蒸汽壓力之上的壓力便會崩潰。 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 wave

37、form 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) illust

38、rates 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). 為了解釋氣穴的影響，考慮到柱塞壓力波時間會體現(xiàn)在圖 3，這就表明空化現(xiàn)象 會發(fā)生在吸力沖程中，當波形中氣穴比例擴大時，高壓尖峰將會接近 0.00025 每秒，通常在源自脈動的復雜的電波中能觀測到氣穴。其中“平方壓力”出現(xiàn)在當蒸汽

39、壓力到達波槽的情況下。顯示氣穴現(xiàn)象的數(shù)據(jù)類型解釋了電波的平方關(guān)閉， 隨之而來的是嚴重的氣穴現(xiàn)象下尖銳的摩擦。 這些取自管道單元 1 的數(shù)據(jù)顯示了高達 600psi 的高壓尖峰。 FIGURE 3. Pump Plunger Pressure-time Wave Showing Cavitation 泵柱塞壓力與時間關(guān)系波形 顯示氣穴現(xiàn)象

40、 FIGURE 4. Cavitation Caused By High Pulsations 高脈動引起的氣穴現(xiàn)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

41、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 f

42、rom the static pressure. Since pulsations greater than 75 psi were always present at the higher speeds, cavitation always occurred. 通過把吸入壓力提高到最大可能值來調(diào)查氣穴現(xiàn)象下靜態(tài)壓力的影響。 單獨靠增加吸入壓力不足以消除氣穴現(xiàn)象。平均增加 200psi 對峰峰值時會有嚴重的脈動。在吸入壓力為 76.5psi 和將近 75psi 零峰值情況下的脈動都會引起氣穴現(xiàn)象。 通過從靜壓力減去脈動負峰從而得到這個值。由于高于 75psi 的脈動通常存在于更高的速

43、度條件下，因此氣穴現(xiàn)象時有發(fā)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

44、 the units. 氣穴現(xiàn)象存在時，實際操作上不能去評估各種影響，例如其他單位，速度，以及 蓄能器設計的影響。 顯然，為了得到單元中有意義的測驗數(shù)據(jù)， 減少高壓震動很有必要。 Acoustical Resonances . 聲學共振 The major succion pulsation -components were at frequencies near ll0 to 150 Hz with pulsation amplitudes of approximately 100 - 150 psi (689 - 1034 kPa) pea

45、k-to-peak, which, when combined 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 whene

46、ver one of the harmonics 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. 主要的吸入震動成分，頻率在近 110-1

47、50Hz，震動振幅接近 100-150psi 對峰值，加上在較低的脈動泵和聲學， 所造成的靜態(tài)壓力降至蒸汽壓力， 毫無疑問，脈動振幅講引起聲學共振。 無論何時一個泵的和聲學的速度超過共振頻率， 聲學共振都將擴大振動。 130Hz 的聲共振是一種吸入管道季波的共振， 并與來自累加到蓄能器的九英尺長的吸管尾端聯(lián)結(jié)。 When an acoustical resonance is encountered in a system, the pressure pulsations can be reduced by eliminating the resonance or b

48、y attenuating the amplitudes through the addition of a resistiveelement, such as an orifice. Therefore, an orifice plate was instaLled at the suction flange in an attempt to attenuate the pulsation amplitudes and possibly move the acoustical natural frequency. A diameter ratio (or

49、ifice diameter to inside diameter of pipe) of approximately 0.4 was used to ensure a significant acoustical effect. When the orifice plate was installed, the pulsations were reduced; however, the reduct:ion was not sufficient to completely eliminate the cavitation. 當系統(tǒng)

50、中遭遇聲學共振， 通過消除共振或是通過增加諸如孔之類的電阻元件來減少振幅便可以減少振幅。于是，在吸力凸緣上安裝一個孔板試圖減少脈動振幅，并可能移除聲學固有頻率。一個大約 0.4 的直徑比（孔口管道內(nèi)直徑）被用來確保重要的聲學效果。 當孔板被安裝后， 脈動即會減少， 但是減少的程度還不足以完全消除氣穴現(xiàn)象。 Interaction With Other Pumps . 與其他泵的相互作用 All the other pumps were shut down and pump I was run to determine if the cavitation was c

51、aused 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 pump

52、 in the manifold system was not a major factor in the cavitation. This is verified by the fact that cavitation 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 location

53、 of the pump in the header was a prime facCor, t:here would have been different pulsation and cavitation characteris tics .To further investigate the interaction of the other pumps, tests were made with pump I on the verge of cavitation and the adjacent unit 2 was swept through the entire

54、 speed range to determine if it affected the speed at which cavitation occurred. This test showed that the adjacent unit: did not influence the cavitation.In an attempt to deterraine whether the acoustical resonance was associated with a piping length from the other units, the suction bloc

55、k valve 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. 所有其他的泵都關(guān)閉后， 泵 1 繼續(xù)運行，通過與其他泵相互作用或是個人管道

56、設計的作用來決定氣穴現(xiàn)象會否發(fā)生。 這些測試表明脈動是由單個泵引起的且其主要原因是聲學共振接近 130 赫茲。此項測試也證明了在多方面系統(tǒng)中泵的位置不是氣穴現(xiàn)象的主要原因。 同一操作條件下， 氣穴現(xiàn)象以完全相同的速度發(fā)生在單元 1 和 3，通過這一事實，此項測試也得以證實。單元 1 和 3 被單元 2 以 32 英尺的距離在二者中間隔開。 如果頭泵的位置是一個主要原因， 那么將會有不同的脈動和氣穴特征。 為了進一步研究與其他泵的相互作用， 進行了相關(guān)測試， 在氣穴邊緣以及與其毗鄰的單元 2 覆蓋了整個速度范圍來確定是否是它影響了氣穴發(fā)生時的速度。 此項測試還表明與其毗鄰的單元并沒有影響到氣穴發(fā)

57、生。 為了摸清聲學共振與其他管道長度是否有關(guān)系， 吸座閥被瞬間壓緊來判斷蓄能器的上流一側(cè)壓力驟降時會否影響到 130 赫茲范圍內(nèi)的共振。座閥上接近 10psi 的壓力下降并無決定性的影響。 Finalng Testing. 最終測試 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 ove

58、r much of the speed range making it possible to study the effect of varying system parameters. The normal procedure 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 o

59、ther pumps, etc.) and then change the pump speed from 190 rpm to 290 rpm. During the speed run, the pulsations 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 s

60、peed 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

61、pump speed. 孔板被安裝后， 且安裝在吸入系統(tǒng)中的氮充電式蓄能器得到最優(yōu)氣體充電后， 氣穴現(xiàn)象將在更大范圍內(nèi)消除， 這就使得研究多變系統(tǒng)指標的作用成為可能。 正常的測試程序是建立一系列的穩(wěn)座條件 （如吸入壓力， 瓶中氣體容量， 氣囊式蓄能 器中的充電壓力，其他泵的速度等等）然后把泵的速度從 190 轉(zhuǎn)切換到 290 轉(zhuǎn)。在此速度運行中， 吸排管道的震動被磁帶錄下用以稍后評估。 這些數(shù)據(jù)表明氣穴的首要原因是在將近 130 到 140 赫茲的興奮是由為 21 泵的額速度通過第三十次諧波系統(tǒng)的聲學固有頻率高的水平脈動。

62、 FIGURE 5. Speed Raster Of Pump Suction Pulsations 泵吸脈動的速度光柵 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 pressur

63、e 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

64、trough. As the speed was further increacedthe cavitation became more severe . 氣穴現(xiàn)象中速度的影響可以在圖 6 中看到，圖 6 給出了速度從 220 到 270 的復雜的壓力波，吸入壓力為 60psi，脈動會速度增長除非有聲學共振。如數(shù)據(jù)顯示，當速度增加到高于 250 轉(zhuǎn)時，脈動增加到一個點， 負壓脈動振幅是蒸汽壓力和波形在槽中變得平穩(wěn)。由于速度變得更快，氣穴現(xiàn)象便會變得更加嚴重。 FIGURE 6. Complex Wave Of Pressure Pulsation Versus

65、 Speed For Suction Pressure Of 60 Psig 壓力脈動對應的復雜波形 60psig 吸入壓力的速度 Static Pressure Effects. 靜態(tài)壓力的影響 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 suctio

66、n 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. 當靜態(tài)吸入壓力達到 90psi 時，脈動振幅便會減少，機組能運行至 280 轉(zhuǎn)，且沒有氣穴（見圖 7）.較高的吸氣壓力似乎抑制了脈動振幅。這些測試的結(jié)果表明，通過增加吸氣壓力到最大可能值， 安裝一個孔板來減少脈動振幅， 并確保蓄能器供電充足，那么，氣穴現(xiàn)象的發(fā)生就可以最小化。