用單極電液伺服閥控制軸向柱塞泵

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1、本科學(xué)生畢業(yè)設(shè)計（論文）附件 附件C：譯文 指導(dǎo)教師評定成績 (五級制)： 指導(dǎo)教師簽字： 附件C：譯文 用單極電液伺服閥控制軸向柱塞泵 愛荷華州立大學(xué)工程研究學(xué)院工程科學(xué)與機械學(xué)系（愛荷華州） 50011 A. Akers 穆爾黑德州立大學(xué)工業(yè)研究部（明尼蘇達州 穆爾黑德）56560 S. J. Lin 【摘要】最優(yōu)控制理論應(yīng)用于一個軸向活塞泵和單級電液伺服閥組合的壓力調(diào)節(jié)器設(shè)計。該控制閥已建模，最優(yōu)控制的規(guī)則也已經(jīng)制定。為了開環(huán)和優(yōu)化控制系

2、統(tǒng)，已經(jīng)獲得了流量階躍輸入的時間響應(yīng)曲線和輸入伺服閥的電流強度。實驗結(jié)果已經(jīng)和那些沒有被作為藍本的斜盤式制動器的供應(yīng)閥門進行比較。該伺服閥控制系統(tǒng)的建模意味著系統(tǒng)的響應(yīng)頻率和壓力峰值的極大提高。 【引言】軸向柱塞泵在航空、工業(yè)、農(nóng)業(yè)系統(tǒng)中都很重要。該泵可以傳送大量的特殊能量，還可以改變能量的流量。對軸向柱塞泵的流量和壓力的控制是通過改變斜盤的角度來實現(xiàn)的。該斜盤驅(qū)動器是由單級或二級的電液伺服閥進行控制的。單級伺服閥是由一個力矩馬達直接連接一個四通滑閥而組成的。閥芯閥由力矩電機定位，由液壓執(zhí)行器指揮控制流向（圖1）。二級伺服閥有一個用于倍增力矩電機輸出的前置放大器，足以克服流體黏附力和由加

3、速度或振動產(chǎn)生的力。插板，噴氣管，閥門和閥芯可作為第一級，而第二級幾乎是普遍的閥芯的類型。 從歷史上看單級伺服閥的穩(wěn)定性和反應(yīng)都優(yōu)于那些使用二級的，但是，自從重量在航天系統(tǒng)中變得特別重要，近期的努力重點放在了完善更輕便的二級伺服閥。然而，工程的緊密公差要求及其他因素導(dǎo)致成本過高，因此，單級伺服閥更可能用于工業(yè)應(yīng)用，因為具有競爭力的價格是必要的。此外，流體動力元件設(shè)計者認為產(chǎn)生相對較大的閥芯力量是很有必要的。一些不可避免的出現(xiàn)在液壓油和有時出現(xiàn)在氣閥座上的力量（約100牛頓），往往會切斷金屬或其它芯片，這點在二級閥閥芯的線軸上是不會出現(xiàn)的。 好幾項致力于研究和改善軸向柱塞泵動態(tài)控制系統(tǒng)的研究

4、已經(jīng)在進行當中。Harpur [1] and Merritt [2] 使用線性擾動分析來研究有微分區(qū)插孔的三通伺服閥和有等面積插孔的四通伺服閥的控制系統(tǒng)。Dreymuller [3] 用勞斯系數(shù)數(shù)組研究軸向柱塞泵的最佳性能。Mack et al. [4] 驗證了給變量泵安裝微機接口用于控制泵的流量和壓力以達到對泵的動作進行補償?shù)目尚行?。最近，Zeiger and Akers [5] 應(yīng)用最優(yōu)控制理論為軸向柱塞泵設(shè)計了一個壓力調(diào)節(jié)器。他們的研究結(jié)果表明，直線性最優(yōu)控制方法沒有為流量干擾提供足夠的壓力強度。不過，增強最優(yōu)控制通過它的流量干擾抵消能量，提供了良好的解決辦法。他們的工作并沒有考慮到使用

5、的伺服閥的類型，而是提出了這樣一個疑問：這樣一個裝置的頻率和阻尼能夠得到什么樣的代表值。相關(guān)伺服閥的短缺使得人們認為對于設(shè)計泵時使用不同伺服閥的影響的全面調(diào)查相當重要。本文介紹了單級伺服閥進行的工作，考慮第一伺服閥型。 在這項工作中，推導(dǎo)出了軸向柱塞泵系統(tǒng)的狀態(tài)方程。此外，Zeiger and Akers [5] 依靠單級伺服閥，并且對壓力時間曲線進行比較從而對泵的斜盤的驅(qū)動建模。 圖1 泵系統(tǒng)的物理模型 一、 動力系統(tǒng)模型 到控制執(zhí)行器的流量連續(xù)性忽略了可壓縮性的影響，表示為 （1） 當流量連續(xù)性的原則是適用于在泵的排放量控制線，我們得到 （2） 關(guān)于

6、斜盤活塞的任何角位置的瞬時扭矩在參考文獻（6）中已得出。通過該模型計算出的力矩的準確性總在實驗值的10%以內(nèi)。研究結(jié)果還表明，扭矩和壓力之間的關(guān)系因為泵的不同而不同，斜盤傾角和斜盤角速度在實際范圍內(nèi)大致呈線性關(guān)系。這種分析使我們能夠編寫在一個線性方程形式扭矩，由于斜盤施加的扭矩是由執(zhí)行機構(gòu)平衡（有彈簧和壓力的力量對他們進行作用）。因此 （3） 該永磁力矩電機用于移動的伺服閥閥芯產(chǎn)生轉(zhuǎn)矩由下式給予 （4） 對轉(zhuǎn)子運用牛頓第二定律，我們得到 （5） 對作用在閥芯上的流體壓力也同樣進行了分析，我們可以把方程（5）寫成 （6） 因為

7、 狀態(tài)變量的分配如下： （7） （8） 兩個控制輸入如下 （9） 方程（2），（3）及（6） - （9）可變成狀態(tài)和輸出方程形式： （10） （11） 該軸向活塞泵控制框圖單級伺服閥如圖2所示 ，它顯示了功能組件是如何聯(lián)接的，并且確定了每個組件的響應(yīng)數(shù)學(xué)方程。 二、最優(yōu)控制設(shè)計 該控制系統(tǒng)的主要目標是設(shè)計一個控制規(guī)則，使輸出（整個泵壓力變化的不同）接近于0，這樣可以流量和其他因素變化的時候保持恒定的壓力。此外，過度的壓力峰值能夠維持在可承受的限度內(nèi)。線性動態(tài)系統(tǒng)的最優(yōu)調(diào)節(jié)器問題表現(xiàn)為包括

8、確定一個矢量控制u（t）的最小化功能 （12） 根據(jù)已知條件 如果是一個真正的對稱，半正定矩陣，是一個真正的對稱，正定矩陣。 最優(yōu)控制法把價值函數(shù)降低到最小值，方程（12）可以被轉(zhuǎn)化為 （13） 當反饋增益矩陣為 (14) 并且當是對稱的，正面的穩(wěn)態(tài)方程定解 （15） 當矩陣和在方程（12）中所代表的相對比較重要的值都被精確確定，把方程（13）中的最優(yōu)控制規(guī)則作為條件借出方程（15），這些值不斷的傳送給數(shù)據(jù)處理器，產(chǎn)生反饋信號對伺服閥進行控制。 最佳閉環(huán)系統(tǒng)，可得到如下 (18) 特征值

9、是矩陣和閉環(huán)系統(tǒng)的極點。該比重可能會取得理想的瞬時變化的反應(yīng)[7]。 圖2 單級伺服閥軸向柱塞泵的結(jié)構(gòu)框圖 三、結(jié)果 開環(huán)系統(tǒng)的初始數(shù)據(jù)，我們可以單獨的使用一個干擾步驟輸入包括。下游量的數(shù)值可以假設(shè)為0.5， 1.0和2，壓力波動，斜盤角，斜盤角速度就可以求解。正如預(yù)期的達到峰值壓力干擾的時間隨著V的增加而增加，隨著初始速率的變化，壓力始終保持不變。當0.01/s用作一個輸入步驟時，穩(wěn)定的斜盤角度被看作與下游量成正比。而且當下游量增長的時候，斜盤也會有一個超過起穩(wěn)態(tài)位置的增長趨勢。當使用干擾電流的時候，初速度的變化是與下游量無關(guān)的。 四、最優(yōu)控制系統(tǒng) 對最優(yōu)控制器的性能已

10、經(jīng)進行了研究，它的計算結(jié)果和參考文獻【5】所展現(xiàn)的一樣，缺乏對現(xiàn)存控制系統(tǒng)進行改善的穩(wěn)健性。因此，它不會被完全的用作為一個控制器。為本文所使用的配置響應(yīng)計算結(jié)果再次證實，此方面還需要增強。這個增強只須將最佳比例控制器轉(zhuǎn)換成比例加積分控制器。這種將有助于增強響應(yīng)時間，超調(diào)量，并減少穩(wěn)態(tài)誤差。本文中使用的這種控制器，如圖（3）所示，可以看出，使用增量積分使這個增強達到了一個相當完美的程度。這個過程導(dǎo)致了一個六指令系統(tǒng)。 該方法用于分析選擇矩陣Q和由產(chǎn)生的不同控制規(guī)則組成的標量、工程量R的影響，然后計算性能指標和策繪了從方程解最優(yōu)控制泵的根位置。為Q矩陣和R選擇合適的值以達到減少穩(wěn)定時間、超調(diào)量和

11、穩(wěn)態(tài)錯誤的系統(tǒng)響應(yīng)的結(jié)果。挑選的值如下： 用來對干擾進行響應(yīng)的數(shù)據(jù)是與一個系列22泵和一個典型的單級電液伺服閥 聯(lián)系的。這些數(shù)據(jù)在表一中給出，其中為流動性能的穩(wěn)定值也一并給出。 圖3 加強的最優(yōu)控制 五、響應(yīng)圖 對于一個階躍的響應(yīng)已經(jīng)表示在圖4至7里面。該系統(tǒng)最佳下游量再次假定為。對于有下游量 和 的次優(yōu)的系統(tǒng)，同樣進行了研究。壓力反應(yīng)圖4所示的是V的較小速度和較小峰值。此外，在參考文獻【5】中可以知道，當泵在最優(yōu)化控制與正確的單級閥模型條件下時，響應(yīng)頻率大約是平時的三倍，而峰值的壓力也同樣因為這個三倍的變化而減少了。 當循環(huán)是封閉的時候，響應(yīng)是相當?shù)目?。從圖6和7

12、可以觀察到V值對斜盤角速度、閥芯移位和峰值速度有非常重要的影響。此外，當下游量減半或變成最優(yōu)值的兩倍時，頻率響應(yīng)也有±20%的變化。 圖8給出了將泵的轉(zhuǎn)速從210rad/s減少到126 rad/s對響應(yīng)的影響。通過對Akers and Zeiger [5]的工作的再次直接比較，可以看出壓力峰值的過沖大幅減 圖5 斜盤角度的時間響應(yīng)（最優(yōu)系統(tǒng)）， 圖4 優(yōu)化系統(tǒng)的壓力——時間響應(yīng)曲線， 圖6 斜盤角速度的時間響應(yīng)（最優(yōu)系統(tǒng)）， 圖7 閥芯位移的時間響應(yīng)（最優(yōu)系統(tǒng)）， 少而頻率卻增加了。參考文獻【5】中提及，該曲線也顯示出這樣一個相同的趨勢：當轉(zhuǎn)速提高時，較小峰值和頻率同樣也會增大

13、。 六、結(jié)論 用一個單級伺服閥以驅(qū)動斜盤并且控制泵壓來建立一個商用的軸向柱塞泵模型是可行的。 開環(huán)系統(tǒng)已經(jīng)開始研究，并且最優(yōu)控制準則也已經(jīng)在那時制定。最優(yōu)閉環(huán)系統(tǒng)的時間響應(yīng)曲線已經(jīng)提出。 閉環(huán)系統(tǒng)、優(yōu)化系統(tǒng)時間響應(yīng)之間的比較同時在開環(huán)系統(tǒng)和一個用工程實踐【5】假定頻率值和斜盤控制執(zhí)行機構(gòu)的阻尼值的系統(tǒng)中進行。在每一個比較當中，單級閥的生產(chǎn)納入了顯著低峰值壓力、更高響應(yīng)頻率的指標以改進其性能。單級伺服閥性能提升的重要原因是因為有準確的建模以進行狀態(tài)變量分析。 圖8 最優(yōu)系統(tǒng)下，泵不同轉(zhuǎn)速的壓力時間響應(yīng)，v=12， 表一 使用的數(shù)據(jù)系列22軸向柱塞泵和一個典型的單級電液伺

14、服閥： = 1000 MPa (油壓縮) （泄露系數(shù)） （量排放系統(tǒng)） （軸轉(zhuǎn)動） （具體的體積位移） (斜盤慣性) （斜盤力矩系數(shù)） （執(zhí)行器彈簧剛度） （活塞球半徑） （斜盤力矩系數(shù)） （斜盤力矩系數(shù)） （流量壓力系數(shù)） （位移執(zhí)行器的控制容積） （活塞領(lǐng)域） （流量增益） （慣性電樞） （電樞支點的距離） （閥芯質(zhì)量） (粘性阻尼系數(shù)) (閥芯阻尼系數(shù)) （力矩電機彎矩） (流通領(lǐng)域梯度) (恒轉(zhuǎn)矩電機) (供應(yīng)壓力) 七、答謝 感謝愛荷華州Sundstrand HydroTransmission公司提供泵的幾何數(shù)據(jù)和一種單級伺服閥

15、的有關(guān)數(shù)據(jù)。此項成果由工程科學(xué)和機械系支持，愛荷華州立大學(xué)工程研究所提供資金和編輯幫助。對于他們的支持非常感謝。 參考文獻： 【1】Harpur, N. F., "Some Design Considerations of Hydraulic Servos of the Jack Type," Proc. Conf. Hydraulic Servomechanism, Vol. 41, I Mech E (1953). 【2】Merritt, H. E., Hydraulic Control Systems, John Wiley & Sons lnc. (1967).

16、 【3】Dreymuller, J., "Pilot-operated and Directly Actuated Pressure Control with Variable Delivery Axial Piston Pumps," Proc. 4th International Fluid Power Symposium, pp. B1-1 to BI-B20 (1975). 【4】Mack, P., et al., "Microcomputer Control of a Variable Displacement Pump," Proc. 40th Nat, Conf. on Flu

17、id Power, Vol. XXXVIII, 55-61 (1984). 【5】Zeiger, G., and Akers, A., "Optimal Control of an Axial Piston Pump," Proc. 7th International Fluid Power Symposium Paper No. 7, 57-64 (1986). 【6】Zeiger, G., and Akers, A., "Torque on the Swashplate of an Axial Piston Pump," ASME Journal of Dynamic Systems

18、Measurement and Control, Vol. 107, No. 3, 220-226 (1985). 【7】D'Azzo, J. J., and Houpis, C. H. Linear Control System Analysis and Design, 2nd edition McGraw-Hill Book Company (1981). CONTROL OF AN AXIAL PISTON PUMP USING A SINGLE-STAGE ELECTROHYDRAULIC SERVOVALVE A. Akers Eng

19、ineering Research Institute and Department of Engineering Science and Mechanics Iowa State University Ames, Iowa 50011 S. J. Lin Department of Industrial Studies Moorhead State University Moorhead, Minnesota 56560 ABSTRACT Optimal control theory is applied to the design of a pressure regulator f

20、or an axial piston pump and single-stage electrohydraulic valve combination. The control valve has been modeled and an optimal control law has been formulated. The time response curves due to a step input in flow rate and in current input to the servovalve have been obtained for the open loop and fo

21、r the optimal control system. The results have been compared to those in which the supply valve to the swashplate actuators was not modeled. Controlled system modeling of the servovalve significantly improves the system's response frequency and pressure peaks. INTRODUCTION Axial piston pumps are

22、 important in aircraft, industrial, and agricultural systems; they can transmit large specific powers and their flow rate can be varied. The control of flow or pressure of axial piston pumps is achieved by changing the swashplate angle. The swashplate actuator is controlled by an electrohydraulic se

23、rvovalve, which may be either single-stage or two-stage. Single-stage servovalves consist of a torque motor directly attached to a four-way spool valve. The spool valve, positioned by the torque motor, directs controlled flow to the hydraulic actuator (Fig. 1). Two-stage servovalves have a hydraulic

24、 preamplifier that multiplies the force output of the torque motor sufficiently to overcome flow and stiction forces and forces resulting from acceleration or vibration. Flapper, jet pipe, and spool valves may be used as a first-stage, while the second-stage is almost universally of the spool type.

25、 Historically, the stability and response of the single-stage servovalves have been superior to those using two stages, but since weight is paramount in aerospace systems, recent effort has focused on perfecting the lighter, two-stage servovalve. However, the close engineering tolerances required an

26、d other factors have led to high costs; thus, it is more likely that single-stage valves will be used for industrial applications where competitive pricing is essential. In addition, designers of fluid power components see a need to generate relatively large spool forces. Such forces (approximately

27、100N), not achievable in the spools of two-stage valves, are required to sever metal or other chips, which inevitably are present in hydraulic oil and which sometimes lodge at the valve seats. Several studies have been conducted in an attempt to investigate and improve the dynamic control systems o

28、f axial piston pumps. Harpur [1] and Merritt [2] used linearized perturbation analysis to investigate a three-way servovalve with differential area jack and a four-way servovalve with equal-area jack control systems. Dreymuller [3] analyzed optimal performance of axial piston pumps by use of a Routh

29、 coefficients array. Mack et al. [4] examined the feasibility of interfacing a microcomputer to a variable displacement pump to control flow rate and provide pressure compensation for the pump action. More recently, Zeiger and Akers [5] applied optimal control theories to the design of a pressure re

30、gulator for an axial piston pump. Their results showed that a straight-linear, optimal control method did not provide adequate pressure stiffness to flow disturbances. However, the augmented optimal control provides a good solution to the pump regulator problem because of its capability to offset fl

31、ow disturbances. Their work did not take into account the type of servovalve to be used; an assumption was made as to what representative values could be obtained for frequency and damping of such a device. The lack of an associated servovalve was considered sufficiently serious to warrant a full in

32、vestigation into the effects of using different servovalve designs with the pump. This paper describes the work conducted with a single-stage servovalve, the first servovalve type considered. In this work, the state equations were derived for the axial piston pump system. In addition, actuation of

33、the swashplate of the pump by means of a single-stage servovalve was modeled, and a comparison was made between the pressure time-response curves obtained and those obtained by Zeiger and Akers [5]. DYNAMICAL MODEL OF THE SYSTEM [2] The flow continuity into the control actuator neglecting the effe

34、ct of compressibility, is expressed as (1) When the principle of flow continuity is applied to the control volume in the discharge line of the pump, we have （2） The instantaneous torque on the swashplate at any angular position of the pistons has been obtained in Ref. [6]. The accuracy

35、of the computed torques from the model is shown to be everywhere within 10% of experimental values. The results also indicate that relationships between torque and pressure differential across the pump, swashplate angle, and angular velocity of the swashplate are approximately linear over the practi

36、cal range. That analysis permits us to write the torque equation in a linearized form, since the torque exerted on the swashplate is balanced by the actuators (which have spring and pressure forces acting on them). Then （3） The permanent-magnet torque motor used to move the spool of the serv

37、ovalve produces a torque given by （4） Applying Newton's second law to the armature, we obtain （5） The stroking flow forces acting on the valve spool have also been analyzed, and we may write Eq. (5) as （6） Where The assigned state variables are as follows: （7）

38、 （8） and two controlled inputs are （9） Equations (2), (3) and (6)-(9) can be put into the state and output equation form: （10） （11） The block diagram of the axial piston pump controlled by a single-stage servovalve is shown in Fig. 2, which shows how the functional components are

39、 connected and the mathematical equations that determine the response of each component. OPTIMAL CONTROL FORMULATION The main objective of the control system is to design a control law that will bring the output (the variation of pressure differential across the pump) close to zero, so that const

40、ant pressure may be maintained while flow rate and other variables change. In addition, over-pressure peaks may be kept within acceptable limits. The optimal regulator problem for linear dynamic systems shown below consists of determining a vector control u(t) that minimizes the functional （12）

41、 Subject to the restrictions where is a real symmetric, positive semi-definite matrix, and is a real symmetric, positive definite matrix. The optimal control law that minimized the cost function, Eq. (12), can be specified as （13） where the feedback gain matrix is (14) and

42、where is the symmetric, positive definite solution of the steady-state Riccati equation （15） Once the matrices and that represent assessment of the relative importance of the various terms in Eq. (12) have been specified, the solution of Eq. (15) specifies the optimal control are law in Eq.

43、(13); these values continuously to the value of the processed and fed back generate control current into the servovalve. The optimal closed-loop system can be obtained as (18) The eigenvalues of are poles of matrices and closed loop system. The weighting may be to obtain desirable transi

44、ent changed response [7]. RESULTS Open Loop System For the first set of data we use a disturbances comprising, separately, a step input of. Values of downstream volume of 0.5, 1.0, and 2were assumed, and the resulting pressure fluctuations, swashplate angle, and swashplate angular velocity were

45、evaluated. As expected, the time to achieve peak pressure disturbance is increased with increasing V, with the initial rate of change of pressure being approximately constant. The steady swashplate angle is seen to be proportional to the downstream volume when a step input of 0.01/s is used, and as

46、downstream volume increases there is an increasing tendency for the swashplate to overshoot its steadystate position. When the disturbance current is applied, the initial rate of change of is independent of downstream volume. Optimally Controlled System The performance of the nonaugmented optimal

47、 controller has been investigated. Its computed results lack robustness for improving the existing control system as shown in Ref. [5]. It therefore would not perform adequately as a controller. The computed results for response for the configuration used in this paper confirmed that once more augme

48、ntation was required. The augmentation simply converts the optimal proportional controller into a proportional plus integral controller. Such an augmentation will aid in response time, overshoot, and reducing steady state error. The controller used in this paper, as shown in Fig. 3, shows where augm

49、entation has been achieved by using the integral of the increment of .This procedure gives rise to a sixth-order system.. The method used for the analysis on the effect of the selection for the Q matrix and scalar quantities R consisted of generating different control laws and then computing the

50、performance index and plotting the root location of the optimally controlled pump from the solution of the Riccati equation. The appropriate values selected for the Q matrix and R result in reduced settling time, overshoot and steady state errors of the system responses. Selected values appear below

51、. The data used for the response to disturbances were those associated with a Series 22 pump and a typical single-stage electrohydraulic servo-valve and are given in Table 1, where the steady values for flow properties are also given. Response Diagrams The responses to a step input of are

52、 shown in Figs. 4 through 7. The downstream volume for the optimal system was again assumed to be . Suboptimal systems having downstream volumes of and were also investigated. The pressure response shown in Fig. 4 is more are smaller rapid and the peaks less for values of V. In addition, it can be

53、 seen that when the pump control is optimized with the correct model of the single-stage valve, then the response frequency is roughly three times that shown in Ref. [5]; the peak pressures are also reduced by a factor of three. —————————————————————————————————— Table 1. Data used for the Serie

54、s 22 axial piston pump and a typical single-stage electrohydraulic servovalve. = 1000 MPa (oil compressibility) —————————————————————————————————— The response is considerably faster when the loop is closed. From Figs. 6 and 7 it is observed that th

55、e value of V has a significant effect upon the angular velocity of the swashplate, the spool displacement, and velocity peak values; in addition, the response frequencies are changed by ±20% when the downstream volume is halved or doubled from the optimal value. On response Figure 8 gives the effec

56、t on the response of reducing the pump rotational speed from 210 to 126 rad/s. Once more a direct comparison with the work of Akers and Zeiger [5] illustrates that the pressure-peak overshoots are much reduced and the frequency is increased. The curves also show a tendency identical to that in Ref.

57、[5] where increase in rotational speed gives rise to a smaller peak response and a higher frequency. CONCLUSIONS It has been possible to model a commercially available, axial piston pump by using a single-stage servovalve to drive the swashplate and so control the pump pressure. The open-loop

58、 system was investigated and an optimal control law was then formulated. Time response curves have been presented for the closedloop optimal system. Comparison between the closed-loop, optimal-system time responses have been made with both the open loop system and with a system where values of freq

59、uency and damping for the swashplate actuator control were assumed by using engineering practice [5]. In each comparison, incorporation of the single-stage valve improves the performance by producing significantly smaller peak pressures and higher frequencies. The principal reason for performance im

60、provement is the fact that the single-stage servovalve has been correctly modeled for inclusion in the state variable analysis. ACKNOWLEDGMENTS Gratitude is expressed to Sundstrand HydroTransmission Company in Ames, Iowa, for furnishing pump geometrical data and relevant data for a typical single-

61、stage servovalve. The work was supported by the Department of Engineering Science and Mechanics, and financial and editorial help was provided by the Engineering Research Institute of Iowa State University. This support is gratefully acknowledged. REFERENCES 1. Harpur, N. F., "Some Design Consider

62、ations of Hydraulic Servos of the Jack Type," Proc. Conf. Hydraulic Servomechanism, Vol. 41, I Mech E (1953). 2. Merritt, H. E., Hydraulic Control Systems, John Wiley & Sons lnc. (1967). 3. Dreymuller, J., "Pilot-operated and Directly Actuated Pressure Control with Variable Delivery Axial Piston P

63、umps," Proc. 4th International Fluid Power Symposium, pp. B1-1 to BI-B20 (1975). 4. Mack, P., et al., "Microcomputer Control of a Variable Displacement Pump," Proc. 40th Nat, Conf. on Fluid Power, Vol. XXXVIII, 55-61 (1984). 5. Zeiger, G., and Akers, A., "Optimal Control of an Axial Piston Pump,"

64、Proc. 7th International Fluid Power Symposium Paper No. 7, 57-64 (1986). 6. Zeiger, G., and Akers, A., "Torque on the Swashplate of an Axial Piston Pump," ASME Journal of Dynamic Systems Measurement and Control, Vol. 107, No. 3, 220-226 (1985). 7. D'Azzo, J. J., and Houpis, C. H. Linear Control Sy

65、stem Analysis and Design, 2nd edition McGraw-Hill Book Company (1981). 指導(dǎo)教師評定成績 (五級制)： 指導(dǎo)教師簽字： 附件C：譯文 壓電陶瓷活塞驅(qū)動液壓泵的發(fā)展 G. W. Woodruff 學(xué)校機械工程學(xué)院，喬治亞理工學(xué)院（喬治亞，亞特蘭大）30332-0405 William S. Oates, Lisa D. buck 和 Christopher S. Lynch 【摘要】 壓電材料在高頻

66、率和大型電場工作時可以產(chǎn)生非常高的功率密度。在馬達中利用次功率密度只會由壓電材料產(chǎn)生很小的行程。行程的整改是為了得到一個高功率密度的設(shè)備。這一點已經(jīng)通過用一個活塞驅(qū)動液壓泵來實現(xiàn)了。更改流體動力也已通過使用止回閥實現(xiàn)。泵的設(shè)計和特性詳情已經(jīng)得到。壓電堆棧的熱力循環(huán)在壓電泵中的應(yīng)用也已闡明。 【引言】 許多應(yīng)用智能材料需要龐大的力量和大的位移。單行程器可以在一個很小的距離內(nèi)產(chǎn)生一個很大的力。超聲波馬達利用共振波來趨勢軸旋轉(zhuǎn)。其結(jié)果是固態(tài)馬達做得非常小，但是，只具有很低的扭矩輸出。分布重復(fù)設(shè)備依靠壓電直接驅(qū)動能力工作時低于共振頻率。這兩個設(shè)備分別是直線電機和液壓泵。本文對液壓泵的發(fā)展和性能進行了詳細的討論。 一、分步重復(fù)執(zhí)行器 分步重復(fù)執(zhí)行器利用壓電的高頻率性能。壓電致動器在每個過程中只能做數(shù)量有限的工作。用與執(zhí)行器阻抗相匹配的負載使工作量達到最大化。在單位時間內(nèi)的能量就是工作量。而這個能量可以通過提高頻率來增加。 該系統(tǒng)利用一個堆棧執(zhí)行器來控制液壓泵里的活塞。系統(tǒng)以提高到頻率的潛力為重點來彌補堆棧的小排量問題。通過在泵中增加一個入口和出口單向閥來糾正流體流動以得到直線運