外文翻譯--對轉(zhuǎn)向裝置液壓伺服系統(tǒng)的動態(tài)仿真

1、<p><b>　　附錄A2</b></p><p>　　對轉(zhuǎn)向裝置液壓伺服系統(tǒng)的動態(tài)仿真</p><p><b>　　摘要</b></p><p>　　基于轉(zhuǎn)向裝置液壓伺服系統(tǒng)的動態(tài)模型可通過力的曲線圖來建立，然后在計算機上對其仿真。通過對已建立的模型在不同情況下的比較，這種比較是通過選擇不同的參數(shù)并通過仿真實

2、現(xiàn)的。在紙上的結(jié)果對于系統(tǒng)的設(shè)計者和分析者來說是有價值的。</p><p>　　關(guān)鍵詞: 組件的選擇和匹配;力學(xué)圖形；AEG轉(zhuǎn)向裝置的液壓伺服系統(tǒng)。</p><p><b>　　1.介紹</b></p><p>　　一個液伺服系統(tǒng)被廣泛地應(yīng)用于板船上的轉(zhuǎn)向裝置。對于一個液壓系統(tǒng)的設(shè)計者來說僅僅知道他提出的系統(tǒng)能夠驅(qū)動負載從一種狀態(tài)到達另一種狀態(tài)

3、或系統(tǒng)是可靠的是不的。他也應(yīng)該知道怎樣選擇元件并正確的相互匹配以減少成本并節(jié)約能耗。這篇文章將通過仿真詳細的介紹液壓伺服系統(tǒng)。</p><p><b>　　系統(tǒng)描述</b></p><p>　　系統(tǒng)可由圖一描述。這個系統(tǒng)被應(yīng)用于“YU CAI”號船上的轉(zhuǎn)向裝置，這只船1970年在德國建造。油的供應(yīng)以及通過控制閥的流動和液壓缸由圖2所示。</p><

4、p>　　Fig. 1 系統(tǒng) Fig. 2 閥及缸中的油液流動</p><p><b>　　2.動態(tài)建模</b></p><p><b>　　(1) 力學(xué)圖形</b></p><p>　　基于力流模型的概念，并根據(jù)鍵合圖形結(jié)構(gòu)的大體脈絡(luò)，系統(tǒng)的鍵合圖模型結(jié)構(gòu)如圖三所示。在這種結(jié)構(gòu)中

5、，閥的泄漏(從來源埠到盡頭)和液壓缸泄漏被忽略。 引動器磨擦片和活塞(包括活塞桿)的慣性Ia已經(jīng)與負載(引動器負荷強行地被加倍)在一起計算。那線路和泵的容量已經(jīng)被增加到過濾器器,被表示成 Cp.</p><p>　　Fig. 3 鍵合圖</p><p><b>　　3.仿真</b></p><p>　　為了要解決模型, 一個程序用ACSL進行。

6、那第四個命令 Runge-Kutta積分法運算法則已經(jīng)在程序中被采用。基于初次的數(shù)值和系數(shù)以及輸入Xv(t)是指定的, 仿真已經(jīng)在計算機上被執(zhí)行。圖4表示輸入Xv(t). 圖5表演負載位移Xm(t)</p><p>　　圖 4 管閥的輸入位移 圖 5 負載位移Xm(t) </p><p><b>　　4.比較并討論</b><

7、/p><p>　　一經(jīng)仿真進行,成象是一件十分簡單的事情。正如像圖5那樣，調(diào)查在特性參數(shù)改變情況下系統(tǒng)的靈敏度。系統(tǒng)改變的動態(tài)變化是可以得到的。二個方法能被采用到變更一個或者較多的參數(shù)而且顯示出所有變數(shù)的時間軌跡。一個將使用外部的運行時間 ACSL 的指令;另外的將在主要的樣板程序中采用一個回路。 采用上述提到的方法,通過運行時間ACSL指令改變Vp和Kv，</p><p>　　在圖6到圖9中

8、二個情況被顯示。圖6和圖7表明如果泵和控制閥做一個完全的匹配,也就是說Vp 和Kv恰當(dāng)?shù)乇幻枋?泵的流量略多于控制閥的額定流量),然后大部份的液體將會被抽到活塞工作;然而,如果泵和控制閥不匹配，也就是說泵的流量比控制閥允許通過的流量大，盡管控制閥保持開著,大部分液體將不得不通過卸荷閥被抽到油缸并且被消耗的能量沒有做任何的有用工作全部轉(zhuǎn)化成了熱量。 這情況在圖8和圖9中顯示出來。</p><p>　　Fig. 6

9、進入控制閥供給液壓缸的流量 Fig. 7 經(jīng)過卸荷閥到油缸的流量 </p><p>　　Fig.8 進入控制閥供給液壓缸的流量 Fig.9 經(jīng)過卸荷閥到油缸的流量 </p><p><b>　　5.結(jié)論</b></p><p>　　系統(tǒng)的主要元件應(yīng)該慎重地選擇尤其選

10、擇泵和控制閥。使用圖表中的過程模型通過描繪仿真結(jié)果能解決問題。圖表中列出的模型和方法也能夠被應(yīng)用于研究轉(zhuǎn)向裝置的動態(tài)伺服系統(tǒng)，可以是系統(tǒng)的設(shè)計也可以是動態(tài)的分析。</p><p><b>　　附錄B1</b></p><p>　　New Concept for Hydrostatic Drive with Control of the Secondary Unit C

11、onnected to the Ring Main System.</p><p>　　Introduction:</p><p>　　This thesis proposes a new concept for "Control of the Secondary Units Connected to the Hydraulic Ring Main System". I

12、n this new concept, we study the ability to connect more than one variable displacement hydraulic motor to the hydraulic ring main system, and all motors can work at variable speeds, and drive different loads. The speed

13、for any hydraulic motor can be changed individually at any time through the operation to any speed required, without any effect on the other. These hydraulic moto</p><p>　　A major part of this work is direct

14、ed towards investigation of the ability of this concept to meet the requirements of the drive concept, the stability of the drives, and their ability to drive the loads maintaining the specific speed.</p><p>

15、;　　Background information:</p><p>　　Hydrostatic Drive is a fluid power technology, which has been used as means of transmitting power. The purpose of hydrostatic transmission is to convert mechanical power i

16、nto hydraulic power, and convert it back into mechanical output power in a form, which matches the speed and torque demands of driven mechanisms or machines.</p><p>　　In drive technology, two parameters are

17、important to the power being transmitted: ?</p><p>　　1) Torque = M (Nm)</p><p>　　2) Speed = n (RPM)</p><p>　　These mechanical parameters correspond respectively to the following par

18、ameters in hydrostatic drives: ? </p><p>　　1) Pressure = P (bar)</p><p>　　2) Flow rate = Q (m3/sec)</p><p>　　In normal power sources e.g. a combustion engine or an electrical motor,

19、 the relation between the mechanical parameters is:</p><p>　　P ??M ????????????????????????????????????????????????????????????????????????????????????????????????????????????????????????????????????????????

20、??(1.1)</p><p><b>　　Where: ?</b></p><p>　　P = Power (KW)</p><p>　　M = Torque (Nm)</p><p>　　??= Angular velocity (rad/sec)</p><p>　　For the same

21、 power source and within efficiency considerations the power transmitted is constant and equal to the maximum power that can be produce by the power source, which is constant.</p><p>　　Then equation (1.1) be

22、comes: ?</p><p>　　Pmax ??M ?????K (1.2)</p><p><b>　　Where: ?</b></p><p>　　K = constant</p><p><b>　　?M ??K/?</

23、b></p><p>　　?M??1/????????????????????????????????????????????????????????????</p><p>　　The relation between torque and speed is vice versa, that is, for any torque there is only one correspo

24、nding speed. In other words, in the case of variable speed, by increasing the speed from minimum to maximum speed, the torque will reduce from maximum to minimum torque for the same power being transmitted,which is the m

25、aximum power produced by the power source, after efficiency considerations.</p><p>　　In hydrostatic drives, the story is different.</p><p>　　The equation (1.1) becomes: ?</p><p>　　d

26、P ??Q??p (1.3)</p><p><b>　　Where: ?</b></p><p>　　dp = pressure difference across the motor</p><p>　　The hydrostatic transmi

27、ssion is capable of maintaining a pre?set high?pressure level in the hydraulic circuit,while changing the flow rate, by using a pressure regulated pump and a variable displacement motor.</p><p>　　Then equati

28、on (1.3) becomes: ?</p><p>　　P ??Q ??dp ??Q??K</p><p><b>　　Where: ?</b></p><p>　　K = constant</p><p>　　?P ; Q (1.4)</p&g

29、t;<p>　　From the equation (1.4), the speed in terms of flow rate (Q) has proportional relation with the power input to the hydrostatic transmission. In other words, in the case of variable speed, by increasing the

30、 speed in terms of flow rate (Q), from minimum to maximum speed, the power input to the hydrostatic transmission will increase according to equation (1.3), while the torque in terms of pressure difference across the moto

31、r (dp), remains constant at the maximum value. The characteristic of hydros</p><p>　　Dependent upon the couplings type of the hydraulic parameters two drive concepts can be defined: ?</p><p>　　1

32、) Drive systems with flow coupling (conventional systems).</p><p>　　2) Drive systems with coupling via the operating pressure (systems with control of the secondary unit).</p><p>　　Drive Systems

33、 With Flow Coupling (Conventional Systems): ?</p><p>　　The simple hydrostatic drives (one output), consist of the primary unit (pump) and the secondary unit (motor), which are inter?related by virtue of the

34、hydraulic flow "Q" in (m3/sec). Fig. (1.1) shows this relationship using a closed circuit drive.</p><p>　　The volumetric flow "Q"(which is determined by the input drive speed "n1&quo

35、t; in rpm and the pump displacement "v1") causes the hydraulic motor to achieve an output speed "n2" dependent upon its displacement "v2".</p><p>　　This system is relatively eff

36、icient because no throttling elements are present.</p><p>　　In heavy engineering applications for multi?output circuits, it is common practice to install hydraulic systems with a central oil supply on a so?c

37、alled "RING MAIN SYSTEM", because the power supply unit can be sized to suit total power required, instead of using individual power packs to supply every hydraulic component. Therefore, a ring main system save

38、s weight and cost. The ring main system operates at a constant pressure by using pressure?regulated pumps and with many actuators connected in </p><p>　　There are the ring main system with two actuators in a

39、n open circuit. It also shows a variable displacement motor controlled by three types of control valves; flow control, directional control and counter balance valve. The maximum flow to this motor is limited by a flow co

40、ntrol valve; below the flow control valve the motor is controlled by a directional control valve, which controls the direction of rotation and throttles the flow even further. If the motors tend to act as generators whil

41、e lowe</p><p>　　Through the throttling and under partially loaded conditions, a part of the power is converted into heat losses. This is in fact a part of the pressure generated by the pumps and is not requi

42、red by the actuator at any flow.</p><p>　　With any change in output torque there is change in pressure drop "dp" across the actuator; this change in pressure drop causes the fluid to compress and t

43、hen expand again. This has an adverse effect on the system stability due to the "HYDRAULIC SPRING". </p><p>　　Therefore it is necessary to increase the control time of the pumps to damp the pressur

44、e build?up and keep the system’s stability under control.</p><p>　　Drive Systems With Pressure Coupling (Systems With Control Of The Secondary Unit): ?</p><p>　　It was necessary to look for anot

45、her drive concept which did not have these disadvantages of the drive systems with flow coupling. This drive concept is called "Drive Systems With Pressure Coupling" also known as "The Secondary Control&qu

46、ot;.</p><p>　　The advantages of the secondary control systems are: ?</p><p>　　1) Parallel operation of a number of actuators without limitation.</p><p>　　2) Energy transfer from the

47、 primary to the secondary units without throttling.</p><p>　　3) Energy recovery for use by other actuators or by returning the energy to the primary unit, againwithout throttling.</p><p>　　4) A

48、constant operating pressure in order to eliminate the influence of the hydraulic spring.</p><p>　　5) The ability to include accumulators at any required point.</p><p>　　The Drive System For Seco

49、ndary Control (The New Concept): ?</p><p>　　There are the hydraulic circuit for the secondary control unit connected to the ring main</p><p>　　system and fitted with a speed control device. The

50、hydraulic circuit consists of: ?</p><p>　　1. Variable displacement hydraulic motor.</p><p>　　2. Fixed displacement gear pump.</p><p>　　When the hydraulic motor starts running under

51、the reset speed pressure (demand pressure), the swash?plate actuator will shift to the maximum opening position. At this time the force acting in the actuator are the reset speed pressure, the spring force, and the speed

52、 control pump pressure which is depend upon the hydraulic motor speed. As the hydraulic motor speed increases, the outlet speed control pump pressure increases, and acts with the swash?plate actuator spring force against

53、 the preset pr</p><p>　　control pump pressure decreases, until the reset speed pressure exceeds the sum of the speed control pump pressure and the swash?plate actuator spring force. Then the swash?plate actu

54、ator starts to move in a direction to increase the hydraulic motor displacement to increase the speed. This action will repeat with decay until the swash?plate actuator stops at the equilibrium position, where the summin

55、g forces acting on the swash?plate actuator equal zero, and the hydraulic motor produces enough t</p><p>　　CONCLUSIONS.</p><p>　　The simulation for the new concept secondary controlled hydraulic

56、 drive showed that the hydraulic drives can work at different individual variable speeds with high accuracy and no speed variation at the steady state condition. By using the new concept we can add to the advantages of t

57、he secondary control the following advantages: ?</p><p>　　1. Connect many hydraulic drives to the ring?main system, all the hydraulic drives work at individual variable speeds.</p><p>　　2. Energ

58、y saving.</p><p>　　3. Low coast.</p><p>　　4. No speed variation at the steady state speeds.</p><p>　　5. Hydraulic controlled drive.</p><p>　　References</p><p

59、>　　1. Christophersen, T.O. The Application Of Secondary Control For A Hydraulically Driven Cargo Pump. Thesis for the MSc degree, University of Bath, 1997.</p><p>　　2. Drexler, P. and Faatz, H. Planning A

60、nd Design Of Hydraulic Power Systems, Mannesmann</p><p>　　Rexroth, Gmbh,Lohr am Main, 1988.</p><p>　　3. Feuser, A.and Kordak R. Hydrostatic Drives With Control Of The Secondary Unit, Mannesmann

61、Rexroth, Gmbh, Lohr am Main, 1989.</p><p>　　4. Gawthrop, P. and Smith, L. Met Modelling: Bond graphs and dynamic systems, Prentice Hall International (UK) Limited, 1996..</p><p>　　5. Stecki, J.

62、S. Garbacik, A. and Szewczyk, K. Hydraulic Control Systems, System Design And</p><p>　　Analysis, department of mechanical engineering, Monash University, Australia, 1997.</p><p>　　6. Stecki, J.

63、S. Aleksandersen, J. Conrad, F. Dransfield, P. and B. T. Kuhnell, B. T. Advances In Hydraulic Control Systems, department of mechanical engineering, Monash University, Australia, 1996.</p><p><b>　　附錄B2

64、</b></p><p>　　DYNAMIC SIMULATION OF THE HYDRAULIC</p><p>　　SERVO SYSTEM FOR AEG STEERING GEAR</p><p><b>　　ABSTRACT</b></p><p>　　A dynamic model of a h

65、ydraulic servo system for steering gears is developed by means of power bond graphs and then simulated on a computer. Based on the built model comparison is made between different situations with selecting different comp

66、onents and matching them by simulation. The findings presented in this paper are valuable for the system designers and</p><p><b>　　analysts.</b></p><p>　　KEY WORDS: Components select

67、ion and match; Power bond graphs; The hydraulic servo system for AEG steering gear</p><p>　　INTRODUCTION</p><p>　　A hydraulic servo system is widely used for steering gears on board ship. It is

68、not sufficient for the designer of a hydraulic system only to know that his proposed system will move the driven load from one state to another and the system is reliable. He should also know how the load will move and i

69、f the system response is stable. And further he should know how to select thecomponents and harmoniously match them to reduce the production cost and working energy consumption. This paper will examin</p><p>

70、;　　SYSTEM DESCRIPTION</p><p>　　The system is illustrated in Fig.1. The system has been employed for the steering gear on the vessel “YU CAI”, which was built in Germany in 1970. The oil supply and its flow t

71、hrough the control valve and the cylinder are shown in Fig.2</p><p>　　Fig. 1 The system Fig. 2 Oil flow in the valve and cylinder</p><p>　　DYNAMIC MODELLING</p><p>　　(1)

72、 Power Bond Graph</p><p>　　Based on the concept of power flow modeling, and according to the general criteria for bond graph structure, the bond graph model for the system is constructed (see Fig. 3). In thi

73、s structure, valve leakage (from the source port to the exhaust) and cylinder leakage are neglected. Actuator friction Rfa and piston (including the rod) inertia Ia have been lumped with the load (the actuator load being

74、 rigidly coupled). The capacitance of the line and the pump have been added to the filter, expressed</p><p>　　Fig. 3 The bond graph</p><p>　　SIMULATION</p><p>　　In order to solve th

75、e model, a program has been developed using ACSL. The fourth order Runge-Kutta integration algorithm has been adopted in the program. With the initial values and coefficients and the input Xv(t) specified, the simulation

76、 has been performed on a computer. Fig.4 shows the input Xv(t). Fig. 5 shows the load displacement Xm(t)</p><p>　　Fig. 4 The input displacement of the valve spool Fig. 5 The load displacement Xm(t)<

77、/p><p>　　COMPARISON AND DISCUSSION</p><p>　　Once the simulation is underway, it is a simple matter to make an excursion, such as Fig.5, to investigate the sensitivity of the system to changes in sp

78、ecific parameters. The dynamic development of all system variables is available. Two methods can be adopted to change one or more parameters and show the timedomain locus of all variables. One is to use the external</

79、p><p>　　runtime commands of ACSL; another is to adopt a loop in the main model program. Adopting the methods above-mentioned, with Vp and Kv changed by means of runtime commands of ACSL, two situations are show

80、n in Fig.6 to Fig.9. Fig.6 and Fig.7 indicate that if the pump and the control valve make a perfect match, that is, Vp and Kv are specified appropriately (the pumpflow rate is slightly more than the rated flow rate of th

81、e control valve), then most of the liquid will be pumped to the actuator to do</p><p>　　flow through, although the control valve stays open, lots of the liquid will have to be pumped to the oil tank through

82、the relief valve and the consumed energy will turn into heat without doing any useful work. This situation is shown in Fig.8 and Fig.9.</p><p>　　Fig. 6 The flowrate into the control Fig. 7 The flowra

83、te through the relief </p><p>　　valve supply port valve to tank</p><p>　　Fig.8 The flowrate into the control Fig.9 The flowrate through the relief<

84、/p><p>　　valve supply port valve to tank</p><p>　　CONCLUSIONS</p><p>　　The main components of the system should be well chosen especially the pump and the con

85、trol valve. Using the model developed in this paper can solve the problem by predicting the results through the simulation. _The model and the methods presented in this paper can be also used to study the hydraulic servo

86、 systems for steering gear both in system design and dynamic analysis.</p><p>　　REFERENCES</p><p>　　[1] Dang Kun, Power Flow Modeling of the Servo System for AEG Steering Gear, Proceedings of th

87、e 3rd Nationwide Chinese Youth Symposium on Mechanical Engineering. Beijing, Mechanical Industry Publisher, 1998,11:135-138 (in Chinese)</p><p>　　[2] Dang Kun, Yin Fen, Dynamic Simulation of a Hydraulic Pump