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STOSIAK 弗羅茨瓦夫大學的技術(shù)，wybrzeze wyspianskiego 25，50-370弗羅茨瓦夫，波蘭。 本文對液壓閥上的外部機械振動的影響。理論分析選定的振動絕緣體的貢獻在液壓閥殼體的振動減少了。報道了初步簡單隔振的實驗測試結(jié)果。 關(guān)鍵詞：機械振動，脈動壓力，液壓閥 1 簡介 液壓系統(tǒng)的主要特點是圍繞一個平均值壓力周期性的變化，通常被稱為壓力波動。其后果是缺乏奈特雷負。該泵的位移分量的循環(huán)操作[ 1 ]或在液壓閥的控制元的自我激勵[ 2 ]因流動液體的作用[ 4 ]或外部的機械振動[ 3，5，6 ]是壓力波動的原因之一。壓力波動引起的單獨的系統(tǒng)組件振動。這有不利的影響，特別是對定位的精度，例如，在一個機床刀具。這也適用于（但到一個較小的程度），是影響固定液壓閥的振動源移動機。一般來說，由一臺機器或設(shè)備的振動傳遞復雜的問題可以分為三個相互關(guān)聯(lián)的類別： .振動源， .振動傳遞路徑， .效應。 振動的最常見的原因是與機器的動作或操作連接的干擾，例如，當一個移動臺移動在不平的表面或當旋轉(zhuǎn)件不平衡在材料加工。另一個主要的振動源驅(qū)動單元，例如內(nèi)燃機工作循環(huán)周期時變特性進行[ 7，8 ]。液壓操作系統(tǒng)也是機械振動源引起的壓力波動和位移泵循環(huán)運行期。由于產(chǎn)生的振動頻率不同，傳輸路徑也不同。不規(guī)則的表面上移動的機器動作導致激發(fā)的0.5–250赫茲的頻率范圍為[ 11–9 ]。后者包括由驅(qū)動產(chǎn)生激勵（燃燒）引擎和位移泵運動學，出現(xiàn)壓力波動在機器的液壓系統(tǒng)。由于流動的空氣阻力的振動是在250–16 000赫茲的頻率范圍內(nèi)，他們是由機器的部件分離氣流引起的。同時流動的工作介質(zhì)的液壓系統(tǒng)產(chǎn)生振動和噪聲。有時發(fā)生氣蝕，產(chǎn)生高頻噪聲。振動所產(chǎn)生的機械傳送產(chǎn)生不同的影響。機械振動，影響機器操作員。組件的系統(tǒng)與該機裝備，特別是液壓元件及系統(tǒng)也受到機械振動。這些組件都需要有良好的動態(tài)特性和具有穩(wěn)定性，定位精度高，運行可靠性，確定性，噪音小?，F(xiàn)代液壓比例閥或者液壓微波暴露于外部的機械振動，特別是因為他們中的干擾力可以量的控制力，這可能會導致很多不良影響，如失穩(wěn)，定位不準確，損壞密封件和增加噪聲[ 12 ]。 2 柔性液壓閥固定 正如上面提到的，為了減少液壓閥的控制元件的振動隔離閥殼似乎從底座的外部機械振動感（例如移動機器或機床振動框架）。對振動的外殼專用夾持座水力分布器的設(shè)計是液壓閥靈活的固定效應分析。后者在其兩側(cè)的彈簧支撐系統(tǒng)與一個已知的線性特性和已知的預變形（圖1）。 圖1 氣門座：1–液壓閥（經(jīng)銷商），2–基座，3–彈簧預變形螺栓，4–彈簧，5–移動夾座 該支架的設(shè)計是這樣的，安裝在閥門的彈簧約束（用一個等效剛度）和移動夾座（2，圖1）把它按照干摩擦模型。在其兩側(cè)，由彈簧支撐的價值。一種液壓系統(tǒng)中的比例分配式4wre 6 e08-12 / 24z4 / M操作，如圖2所示。 圖2 液壓系統(tǒng)的組成方案：將調(diào)查1–給水泵，2–溢流閥，3–調(diào)查的組成部分，4–調(diào)節(jié)節(jié)流閥 一二質(zhì)量系統(tǒng)的模型的比例分配在液壓系統(tǒng)如圖2所示，可以通過以下系統(tǒng)的四個方程表示： 第四個方程描述作用在認為情況下閥殼的力量。進一步對該方程將被修改以描述該隔振元件的特性。一些簡化的假設(shè)，方程（1）： 工作液不改變其性質(zhì)， 庫侖摩擦忽略了對閥芯套內(nèi)定向控制閥， 庫侖摩擦是閥體與閥座之間的合作， 彈簧特性是線性的和剛度系數(shù)C描述， 液壓系統(tǒng)的描述是基于集中參數(shù)模型， 該模型不代表管閥體振動的影響。 一個數(shù)值的溶液中形成的“傳遞函數(shù)”， 在閥殼體振動加速度幅值A(chǔ)2激勵振動加速度振幅A0比，如圖3所示。 圖3 比例分配器殼體振動加速度幅值A(chǔ)2相對激振加速度振幅A0 f = 10–60赫茲 對模擬結(jié)果的分析表明，在約20赫茲的頻率振動幅度相當大的增益。這是由于共振自振閥達4.5公斤，持有人的等效剛度的彈簧質(zhì)量86 000 N /米。因此在配器殼體振動的振幅增益OB曾在10–30赫茲的范圍內(nèi)（無效的隔振）。 不同形式的絕緣元件可以假定。一個準零剛度振動絕緣體的引入大大有助于閥門殼體的振動最小化。與準零剛度隔振器的理想的特性是由以下方程[ 13 ]： c1w，C2W–分別主彈簧和補償彈簧的剛度，∝H–角的初始，側(cè)臂軸Y原來的傾向，P1H, P2H–在位置初始彈簧張力H[N]， 在這樣一個振動激發(fā)方向絕緣子總剛度（外部機械振動的方向）是： 因此，模型的第四個方程（1）可以寫為： 模型示例解決方案（1）補充方程（4）是在激勵頻率f = 10–60赫茲以下的數(shù)字顯示。 對模擬結(jié)果的分析表明，由于振動的使用準零剛度閥殼體的振動可以做出降低絕緣子。不過，由于其尺寸絕緣體不能用在小空間。因此，材料具有良好的隔振性能，適合在小空間使用上應尋求?？磥?，特殊墊上安裝液壓閥可以滿足要求。這種材料也應耐液壓油和極端的環(huán)境溫度。 圖5 比例分配器殼體振動加速度幅值A(chǔ)2相對為了激勵振動加速度振幅A0 f = 10–60赫茲 圖6 比例分配器殼體振動加速度幅度A2相對激振加速度振幅A0 f = 10–60赫茲 圖5和6的數(shù)字顯示，這樣一個非線性隔振特性可以選擇，絕緣將在整個考慮激發(fā)頻率范圍內(nèi)有效。 對閥的機械振動的影響這個問題用理論和實驗的方式來考慮。理論上的考慮，基于數(shù)值根據(jù)數(shù)學模型計算。一些理論思考的實驗進行了測試使用測試站（液壓仿真轉(zhuǎn)臺，閥座，彈簧套）。 3 實驗測試 試驗臺上，使機械振動特征的一種規(guī)定的頻率產(chǎn)生了實驗驗證了理論分析的結(jié)果和結(jié)論。研究了閥–曼內(nèi)斯曼力士樂比例分配式4wre 6 e08-12 / 24z4／m–固定在支架安裝在試驗臺和子遭外部機械振動（圖1）。測試是在沒有連接到閥管時進行的。一個線性的靜液壓驅(qū)動模擬器，能夠產(chǎn)生高達100赫茲的振動，是外部的機械振動源。對線性靜液壓驅(qū)動模擬器的主要成分是伺服閥控制液壓缸。該模擬器由三個主要部分：液壓部分，控制部分和控制軟件。模擬表的位移是由位移傳感器和加速度控制是由加速度控制。對仿真轉(zhuǎn)臺測試閥的安裝。模擬電控制信號是由外部諧波信號發(fā)生器的支持。比例分配器放置在專用架雙側(cè)支撐彈簧（有兩個彈簧并聯(lián)在每邊）。初步的測試，用一個等效的彈簧進行了86 000 N / m和2毫米的預變形剛度。外激勵參數(shù)如表2所示。 圖1 比例分配器放置在特殊的支架和兩側(cè)支撐彈簧，在測試過程中 表2 作用于測試液壓分配器的振動振幅 圖8 顯示了一個整體的閥門振動圖的外部激勵，即比例分配器殼體加速度幅值A(chǔ)2激發(fā)振動振幅A0與25–60赫茲的頻率比。 圖8 比例分配器殼體振動加速度幅值A(chǔ)2相對激振加速度振幅A0 f = 25–60赫茲 4 結(jié)論 它已被證明是一個機床和移動設(shè)備的普遍裝備液壓閥振動裝置。絕緣子的振動為特征在一定的外部振動頻率在閥殼體振動加速度振幅降低線性結(jié)果彈簧形式的運用，但它可能有利于共振頻率。在圖3和圖8顯示的結(jié)果比較，模型和測試之間的差異并不很大35–60赫茲的頻率范圍。由于具有非線性特性的閥殼體振動加速度幅值進行幾十%降低隔振裝置的使用：通過與準零剛度隔振器的90%和80%左右的隔振器的剛度或阻尼是位移或速度的第二功率成正比。在閥殼體振動的減少將導致在滑閥減少振動，尤其是共振范圍。在這樣的應用振動絕緣體也應滿足其他的標準，如：耐環(huán)境溫度變化，耐液壓流體，和幾何尺寸小。因此，除了具有良好的理化特性，振動絕緣體，應該有一個標準化的設(shè)計，適合于液壓閥的典型連接板。 參考文獻： [1] Lisowski E., Szewczyk K.: 多活塞軸流泵的理論確定交貨的波動（波蘭），Sterowanie i Napd Hydrauliczny, No. 1, 1984, pp. 3–6。 [2] Kudma Z.: 對減壓閥和液壓系統(tǒng)的自由振動頻率（波蘭），Sterowanie i Napd Hydrauliczny, No. 3, 1990, pp. 27–30。 [3] Amini A., Owen I.: 減壓閥的噪聲與振動問題的一個可行的解決方案，熱流體科學實驗，No. 10, 1995, pp. 136–141。 [4] Misra A., Behdinan K., Cleghorn W.L.: 由于結(jié)構(gòu)相互作用的流體控制閥自激振動，流體與結(jié)構(gòu)雜志，Vol. 16, No. 5, 2002, pp. 649–665。 [5] Stosiak M.: 對液壓閥控制元件的基礎(chǔ)上的低頻機械振動的影響（波蘭），[in:] Rozwój maszyn i urzadzen hydraulicznych, Edit. Wacaw Kollek, Wrocaw, Wydaw. Wroc. Rady FSNT NOT, Vol. 11,No. 2–3, 2006, pp. 83–94。 [6] Stosiak M.: 液壓系統(tǒng)中的壓力脈動對地面機械振動的影響（波蘭），Hydraulika i Pneumatyka, No. 3, 2006, pp. 5–8。 [7] Engel Z.: 對振動和噪聲的環(huán)境保護（波蘭），Wy-dawnictwo Naukowe PWN, Warsaw, 2001。 [8] Leea E.C., Nianb C.Y., Tarng Y.S.: 車削加工中對振動動力吸振器的設(shè)計，材料處理技術(shù)雜志，Vol. 108, 2001,pp. 278–285。 [9] Grajnert J.: 振動絕緣的機械和車輛（波蘭），Oficyna Wy-dawnicza Politechniki Wrocawskiej, Wrocaw, 1997。 [10] Pytlik A.: 在機械外殼部分液壓系統(tǒng)振動（波蘭），Napdy i Sterowanie, Vol. 10, No. 4, 2008, pp. 121–130。 [11] Krylov V., Pickup S., McNuff J.: 從重型軍用車輛的地面振動光譜的計算，聲音與振動雜志，Vol. 329, No. 115, 2010, pp. 3020–3029。 ARCHIVES OF CIVIL AND MECHANICAL ENGINEERINGVol. XI2011No. 1Vibration insulation of hydraulic system control componentsM. STOSIAKWroc?aw University of Technology, Wybrze?e Wyspia?skiego 25, 50-370 Wroc?aw, Poland.This paper deals with the effects of external mechanical vibrations on hydraulic valves. A theoretical analy-sis of the contribution of selected vibration insulators to a reduction in hydraulic valve housing vibrations wascarried out. The results of preliminary experimental tests of simple vibration insulators are reported.Keywords: mechanical vibrations, pressure fluctuations, hydraulic valve1. IntroductionMajor features of hydraulic systems are periodic changes of pressure around an aver-age value, commonly referred to as pressure fluctuations. Their consequences are defi-nitely negative. The cyclic operation of the pump s displacement components 1 or theself-excitation of the control components in hydraulic valves 2 due to the action of theflowing liquid 4 or to external mechanical vibrations 3, 5, 6 are among the causes ofpressure fluctuations. Pressure fluctuations cause the individual system components tovibrate. This has an adverse effect, particularly on the precision of positioning of, for ex-ample, the cutting tool in a machine tool. This also applies (although to a smaller degree)to mobile machines which are the source of vibrations affecting the rigidly fixed hydraulicvalves. Generally, the complex problem of the transmission of vibrations by a machine ora piece of equipment can be divided into three interconnected categories:? vibration sources,? vibration transmission paths,? effects.The most frequent cause of vibrations are disturbances connected with the motionor operation of the machine, for example when a mobile machine moves on an unevensurface or when the rotating parts are unbalanced during material machining. Anothermajor vibration source are drive units, for example a combustion engine performinga periodic variable-characteristic work cycle 7, 8. An operating hydraulic system isalso a source of mechanical vibrations caused mainly by pressure surges and the peri-odic operation of the displacement pump. Since the generated vibrations have differentfrequencies the paths of their transmission are also different. The irregularities of thesurface on which a mobile machine moves cause excitations in a frequency range of0.5 250 Hz 9 11. The latter includes excitations generated by the driving (combus-tion) engine and the displacement pump kinematics, manifesting themselves in pres-M. STOSIAK238sure fluctuations in the machine s hydraulic system. The vibrations due to the resis-tance of flowing air are in a frequency range of 250 16 000 Hz and they are caused byairflow separation from the machine s components. Also the flow of the working me-dium in the hydraulic system causes vibration and noise. Sometimes cavitation occurs,generating high-frequency noise. The vibrations generated and transmitted by a ma-chine produce various effects. Mechanical vibrations affect the machine operator. Thecomponents of the systems with which the machine is equipped, particularly hydrauliccomponents and systems are also subject to mechanical vibrations. Such componentsare required to have good dynamic properties and to be characterized by stability,positioning precision, operating reliability and certainty and little noisiness. Modernproportional hydraulic valves or hydraulic microvalves are particularly exposed toexternal mechanical vibrations since the disturbing forces in them can amount to thecontrolling forces, which may lead to many adverse effects, such as stability loss, po-sitioning inaccuracy, damage to seals and increased noisiness 12.2. Flexible fixing of hydraulic valveAs mentioned above, in order to minimize the vibration of the hydraulic valve scontrol element it seems sensible to isolate the valve housing from the external me-chanical vibrations of the base (for example the vibrating frame of a mobile machineor a machine tool). For the analysis of the effect of the flexible fixing of a hydraulicvalve on the vibration of its housing a special clamping holder for the hydraulic dis-tributor was designed. The latter is on its two sides supported by a system of springswith a known linear characteristic and a known pre-deflection (Figure 1).Fig. 1. Valve holder: 1 hydraulic valve (distributor), 2 holder base,3 spring pre-deflection bolts, 4 springs, 5 securing catchesThe design of the holder is such that the valve mounted in it is constrained bysprings (with an equivalent stiffness) and it moves on the holder base (2 in Figure 1)rubbing against it in accordance with the dry friction model. On its two sides the valveVibration insulation of hydraulic system control components239is supported by springs. A scheme of the hydraulic system in which the proportionaldistributor type 4WRE 6 E08-12/24Z4/M operates is shown in Figure 2.TPAB413w = w0 sin(2 ? f t)Fig. 2. Scheme of hydraulic system incorporating investigated component: 1 feed pump,2 relief valve, 3 investigated component, 4 adjustable throttle valveFor a two-mass system the model of the proportional distributor operating in thehydraulic system shown in Figure 2 can be represented by the following system offour equations:?. 01, 02, 025 . 1,2172. 02222022212112122212211111111212121121212111gimwXsingwXsingwXlHgmwXcXXkXXcXmpACpcpcpaQpcpappxxXsQFXXcppxxXsXXhldXmzaqkkppkpmpspMmpst? ? ? (1)M. STOSIAK240The fourth equation describes the forces acting on the valve housing in the consid-ered case. Further on this equation will be modified to describe the characteristics ofthe proposed vibration insulation elements. Some simplifying assumptions to Equa-tions (1):? working liquid does not change its properties,? Coulomb friction is neglected in pair: spool-muff inside directional control valve,? Coulomb friction represents cooperation between valve body and valve holder,? after play (between valve body and securing catches) is cancelled Coulombfriction represents cooperation between valve body and securing catches,? springs characteristics are linear and described by stiffness coefficient c,? description of hydraulic system is based on concentrated parameter model,? the model does not represent influence pipes on valve body vibrations.List of major symbols:SymbolParameterDimension in SIap1leakage coefficientm4s/kgAathrottle valve gap aream2c1equivalent stiffness of valve centring springsN/mczequivalent stiffness of springs fixing valve in holderN/mCq1throttle valve flow ratio dtpiston diametermffrequencyHzgEarth s accelerationm/s2hvalve-sleeve pair gap thicknessmHHeaviside step function k1, k2damping in respectively valve-sleeve pair and housing-holder pairNs/mlpiston lengthml0gap of valve body and securing catchesmm1mass of piston valve and 1/3 of spring masskgm2mass of distributor housingkgp1pressure before distributorPap2pressure after distributorPapzsink line pressurePa?p2throttle valve pressure dropPassmaximum gap widthmttimeswexcitation vibration amplitudemQptheoretical pump deliverym3/sxmgap lengthmxpmutual shift of valve and housing edgesmX1displacement of piston valvemX2displacement of distributor housingm?2coefficient of friction of valve housing against securing catches ?icoefficient of friction of valve housing against holder base ?working liquid densitykg/m3?angular frequencyrad/sVibration insulation of hydraulic system control components241Model (1) also takes into account the interaction between the valve housing and se-curing catches 5 (Figure 1). A numerical solution in the form of a “ transmission func-tion” , understood as a ratio of valve housing vibration acceleration amplitude a2 toexcitation vibration acceleration amplitude a0, is shown in Figure 3.00,511,522,533,544,51015202530354045505560f Hza2/a0 -Fig. 3. Proportional distributor housing vibration acceleration amplitude a2 relativeto excitation vibration acceleration amplitude a0 for f = 1060 HzAn analysis of the simulation results shows a considerable gain in housing vibra-tion amplitude at a frequency of about 20 Hz. This is due to resonance since the massof the vibrating valve amounts to about 4.5 kg and the equivalent stiffness of the holdersprings is 86 000 N/m. Hence a gain in distributor housing vibration amplitude is ob-served in the range of 10 30 Hz (ineffective vibration insulation).This means that valve insulation which will widen the insulation zone and reducethe resonance zone should be proposed. The black-box approach (Figure 4) was adoptedto solve the problem.Different forms of the insulating element can be assumed. The introduction of a vi-bration insulator with quasi-zero stiffness significantly contributes to the minimizationof valve housing vibrations. The ideal characteristic of the vibration insulator withquasi-zero stiffness is described by the following Equation 13:,cos)(2)2(sin)(222222111HHHwHwwHHwHlxxlcPxcclcPxF? (2)M. STOSIAK242where:c1w, c2w stiffness of respectively the main spring and the compensation spring,?H angle of initial, original inclination of the side arm to axis y,P1H, P2H initial spring tensions in position ?H N,lH length of the side arm in position ?H.m2m1?111cos?txX?222cos?txX?tww?cos0c1k1l0?Fig. 4. Black-box approach to valve vibration insulationThe total stiffness of such a vibration insulator in the excitation direction (the di-rection of the external mechanical vibration) is:.coscoscos)(22)(222222222221?HHHHHHHwHwwlxllxlcPccxc? (3)Thus the fourth equation of model (1) can be written as:?. 0coscoscos)(2222222222222222112112122?wXlXllXlcPccXXkXXcXmHHHHHHHwHww? ? (4)Exemplary solutions of model (1) supplemented with Equation (4) are shown in thefigures below for excitation frequency f = 10 60 Hz.An analysis of the simulation results shows that thanks to the use of the vibrationinsulator with quasi-zero stiffness the vibration of the valve housing can be considera-bly reduced. However, because of its dimensions such an insulator cannot be used insmall spaces. Therefore materials with good vibration insulation properties and suit-able for the use in small spaces should be sought. It seems that special pads (mats) formounting hydraulic valves on them could meet the requirements. Such materials shouldalso be resistant to hydraulic fluids and extreme ambient temperatures. Using theVibration insulation of hydraulic system control components243black-box approach one can select an insulator material characteristic ensuring effec-tive vibration insulation in a wide excitation range.00,050,10,150,20,250,31015202530354045505560f Hza2/a0 -Fig. 5. Proportional distributor housing vibration acceleration amplitude a2 relativeto excitation vibration acceleration amplitude a0 for f = 1060 Hz00,050,10,150,20,250,30,351015202530354045505560f Hza2/a0 -Fig. 6. Proportional distributor housing vibration acceleration amplitude a2 relativeto excitation vibration acceleration amplitude a0 for f = 1060 HzM. STOSIAK244The results of the application of a vibration insulator with characteristic xkxc?222and c2 = 20 000 N/m and k2 = 50 Ns/m are shown in Figure 6. In this case, the fourthequation of model (1) should be supplemented with a nonlinear vibration insulatorcharacteristic.When a vibration insulator with a nonlinear damping characteristic (k2 = 250 Ns/m)and linear stiffness (c2 = 20 000 N/m) 222xkxc? is used to insulate base vibrationsthe valve housing vibrations are as shown in Figure 7.00,511,522,51015202530354045505560f Hza2/a0 -Fig. 7. Proportional distributor housing vibration acceleration amplitude a2 relativeto excitation vibration acceleration amplitude a0 for f = 10 60 HzFigures 5 and 6 show that such a nonlinear vibration insulator characteristic can beselected that the insulation will be effective in the whole considered excitation fre-quency range.The problem of influence of mechanical vibrations on valve was considered in theo-retical and experimental way. Theoretical considerations were based on numericalcalculations according to mathematical model. For some theoretical considerationsexperimental tests were done using test stand (hydraulic simulator, valve holder, springset).3. Experimental testsA test rig enabling the generation of mechanical vibrations characterized by a pre-scribed frequency was built to experimentally verify the theoretical results and conclu-Vibration insulation of hydraulic system control components245sions. The investigated valve Mannesmann-Rexroth proportional distributor type4WRE 6 E08-12/24Z4/M fixed in the holder was mounted on the test rig and sub-jected to external mechanical vibrations (photo 1). Tests were done without pipesconnected to valve.A linear hydrostatic drive simulator Hydropax ZY25 made by Mannesmann-Rexroth,capable of generating vibrations up to 100 Hz, was the source of external mechanicalvibrations. Main component of simulator of linear hydrostatic drive is servo valve whichcontrols hydraulic cylinder. The simulator consists three main parts: hydraulic part,control part and control software. Displacement of simulator table is controlled bydisplacement transducer and its acceleration is controlled by accelerometer. On simu-lator table the tested valve was mounted. Electrical control signal for simulator wassupported by external harmonic signal generator. The simulator is described in moredetail in 4. The proportional distributor was placed in the special holder andbilaterally supported with springs (there were two springs connected in parallel oneach of the sides). Preliminary tests were carried out for springs with an equivalentstiffness of 86 000 N/m and a pre-deflection of 2 mm. The external excitation pa-rameters are shown in Table 2.Photo 1. Proportional distributor placed in special holder and bilaterally supported with springs, during testingTable 2. Amplitude of vibrations acting on tested hydraulic distributorf Hzw0 m300.000483350.000406400.000366450.000269500.000214550.000145600.0000522Figure 8 shows an overall valve vibration diagram for the external excitation, i.e.a ratio of proportional distributor housing acceleration amplitude a2 to excitation vi-bration amplitude a0 versus a frequency of 25 60 Hz.M. STOSIAK246It appears from the diagram shown in Figure 8 that for a system of springs withequivalent stiffness cz = 86 000 N/m and a proportional distributor with a mass of4.5 kg the vibration insulation is effective (transmission function a2/a0 1) in thegiven external vibration frequency range. As a result of the insulation, the distributorhousing vibration amplitude and the distributor slide-valve vibration amplitude de-crease 5. Consequently, the amplitude of the pressure fluctuations due to the excita-tion of distributor slide-valve vibrations also decreases. However, in the case of sosimple vibration insulation, resonance may be generated at external vibration frequen-cies other than the ones used in the test. Therefore, as Figures 5 and 6 indicate, a vi-bration insulation element with other properties and characteristics, e.g. with nonlinearstiffness and with damping, should be used.00,20,40,60,811,22530354045505560f Hza2/a0 -Fig. 8. Proportional distributor housing vibration acceleration amplitude a2 relativeto excitation vibration acceleration amplitude a0 for f = 2560 Hz4. ConclusionIt has been shown that there is a need to reduce the vibration of the hydraulicvalves with which machine tools and mobile machines are commonly equipped. Theuse of vibration insulators in the form of springs whose characteristics are linearresults in a reduction in valve housing vibration acceleration amplitude at certainexternal vibration frequencies, but it may be conducive to resonance at other fre-quencies. Comparison of results presented on Figure 3 and Figure 8 shows, that dif-ferences between model and test are not great for frequency range 35 60 Hz. Thebiggest differences are observed in resonant area (25 Hz). It follows from the pre-sented cases of vibration insulation (Figures 5 8) that materials with linear charac-Vibration insulation of hydraulic system control components247teristics should be used in order to extend the range of effective vibration insulation.Thanks to the use of a vibration insulator with a nonlinear characteristic the valvehousing vibration acceleration amplitude was reduced by a few tens of percent: byover 90% for the vibration insulator with quasi-zero stiffness and by about 80% forthe vibration insulator whose stiffness or damping was proportional to displacementor velocity to the second power. A reduction in valve housing vibration will lead toa reduction in slide-valve vibration, particularly in the resonant vibration range. Asa result, the pressure fluctuations and the emitted noise (particularly in a low fre-quency range) will decrease and the precision of the motions of the hydraulic receiv-ers will increase. Vibration insulators in such applications should also satisfy othercriteria, such as: resistance to changes in ambient temperature, resistance to hydrau-lic fluids, and small geometric dimensions. Therefore, besides having proper phys-icochemical properties, a vibration insulator should have a standardized design suit-able for typical connection plates for hydraulic valves.References1 Lisowski E., Szewczyk K.: Theoretical determination of multi-piston axial-flow pumpdelivery fluctuations (in Polish), Sterowanie i Nap?d Hydrauliczny, No. 1, 1984, pp. 3 6.2 Kud?ma Z.: Frequency of the free vibration of a relief valve and a hydraulic system (inPolish), Sterowanie i Nap?d Hydrauliczny, No. 3, 1990, pp. 27 30.3 Amini A., Owen I.: A practical solution to the problem of noise and vibration in a pres-sure-reducing valve, Experimental Thermal and Fluid Science, No. 10, 1995, pp. 136 141.4 Misra A., Behdinan K., Cleghorn W.L.: Self-excited vibration of a control valve due tofluid-structure interaction, Journal of Fluids and Structures, Vol. 16, No. 5, 2002, pp. 649665.5 Stosiak M.: The effect of the low-frequency mechanical vibrations of the base on the con-trol component of the hydraulic valve (in Polish), in: Rozwj maszyn i urz?dze? hydrau-licznych, Edit. 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Rady FSNT NOT, Vol. 11,No. 2 3, 2006, pp. 83 94.6 Stosiak M.: An influence of mechanical vibrations of ground for pressure pulsation inhydraulic system (in Polish), Hydraulika i Pneumatyka, No. 3, 2006, pp. 5 8.7 Engel Z.: Protection of the environment against vibrations and noise (in Polish), Wy-dawnictwo Naukowe PWN, Warsaw, 2001.8 Leea E.C., Nianb C.Y., Tarng Y.S.: Design of a dynamic vibration absorber against vibra-tions in turning operations, Journal of Materials Processing Technology, Vol. 108, 2001,pp. 278 285.9 Grajnert J.: Vibration insulation in machines and vehicles (in Polish), Oficyna Wy-dawnicza Politechniki Wroc?awskiej, Wroc?aw, 1997.10 Pytlik A.: Vibrations in hydraulic systems of mechanized casing section (in Polish),Nap?dy i Sterowanie, Vol. 10, No. 4, 2008, pp. 121 130.11 Krylov V., Pickup S., McNuff J.: Calculation of ground vibration spectra from heavymilitary vehicles, Journal of Sound and Vibration, Vol. 329, No. 115, 2010, pp. 30203029.M. STOSIAK24812 Kollek W., Kud?ma Z., Maga K., Stosiak M.: Disturbances in the operation of propor-tionally controlled hydrostatic systems (in Polish), Nap?dy i Sterowanie, Vol. 6, No. 5,2004, pp. 53 57.13 Joint publication edited by ?wider J., Kazimierczak J.: Computer-aided design of machinevibration and noise reducing systems (in Polish), WNT, Warsaw, 2001.Wibroizolacja elementw steruj?cych uk?adw hydraulicznychW artykule skupiono si? na problemie oddzia?ywania zewn?trznych drga? mechanicznychna zawory hydrauliczne. Omwiono skutki tych oddzia?ywa?. Przeprowadzono analiz? teore-tyczn? wp?ywu charakterystyki wybranych izolatorw na redukcj? drga? korpusu zaworu hy-draulicznego. Przedstawiono wst?pne badania eksperymentalne dla prostych przyk?adw izo-latorw.