外文翻译--对转向装置液压伺服系统的动态仿真

1 附录 A2 对转向装置液压伺服系统的动态仿真 摘要 基于转向装置液压伺服系统的动态模型可通过力的曲线图来建立，然后在计算机上对其仿真。通过对已建立的模型在不同情况下的比较，这种比较是通过选择不同的参数并通过仿真实现的。在纸上的结果对于系统的设计者和分析者来说是有价值的。 关键词 : 组件的选择和匹配 ；力学图形； AEG 转向装置的液压伺服系统。 1.介绍 一个液伺服系统被广泛地应用于板船上的转向装置。对于一个液压系统的设计者来说仅仅知道他提出的系统能够驱动负载从一种状态到达另一种状态或系统是可靠的是不的。他也应该知 道怎样选择元件并正确的相互匹配以减少成本并节约能耗。这篇文章将通过仿真详细的介绍液压伺服系统。 系统描述 系统可由图一描述。这个系统被应用于 “YU CAI”号船上的转向装置，这只船 1970年在德国建造。油的供应以及通过控制阀的流动和液压缸由图 2所示。 Fig. 1 系统 Fig. 2 阀及缸中的油液流动 2.动态建模 (1) 力学图形 基于力流模型的概念，并根据键合图形结构的大体脉络，系统的键合图模型结构如图三所示。在这种结构中，阀的泄漏 (从来源埠到尽头 )和液压缸 泄漏被忽略。 引动器磨擦片和活塞 (包括活塞杆 )的惯性 Ia已经与负载 (引动器负荷强行地被加倍 )在一起计算。那线路和泵的容量已经被增加到过滤器器 ,被表示成 Cp. 2 Fig. 3 键合图 3.仿真 为了要解决模型 , 一个程序用 ACSL进行。 那第四个命令 Runge-Kutta积分法运算法则已经在程序中被采用。基于初次的数值和系数以及输入 Xv(t)是指定的 , 仿真已经在计算机上被执行。图 4表示输入Xv(t). 图 5表演负载位移 Xm(t) 图 4 管阀的输入位移 图 5 负载位移 Xm(t) 4.比较并讨论 一经仿真进行 ,成象是一件十分简单的事情。正如像图 5那样，调查在特性参数改变情况下系统的灵敏度。系统改变的动态变化是可以得到的。二个方法能被采用到变更一个或者较多的参数而且显示出所有变数的时间轨迹。一个将使用外部的运行时间 ACSL 的指令 ；另外的将在主要的样板程序中采用一个回路。 采用上述提到的方法 ,通过运行时间 ACSL指令改变 Vp和 Kv， 在图 6到图 9中二个情况被显示。图 6和图 7表明如果泵和控制阀做一个完全的匹配 ,也就是说 Vp 和Kv恰当地被描述 (泵的流量略多于控制 阀的额定流量 ),然后大部份的液体将会被抽到活塞工作 ；然而 ,如果泵和控制阀不匹配，也就是说泵的流量比控制阀允许通过的流量大，尽管控制阀保持开着 ,大部分 3 液体将不得不通过卸荷阀被抽到油缸并且被消耗的能量没有做任何的有用工作全部转化成了热量。 这情况在图 8和图 9中显示出来。 Fig. 6 进入控制阀供给液压缸的流量 Fig. 7 经过卸荷阀到油缸的流量 Fig.8 进入控制阀供给液压缸的流量 Fig.9 经过卸荷阀到 油缸的流量 5.结论 系统的主要元件应该慎重地选择尤其选择泵和控制阀。使用图表中的过程模型通过描绘仿真结果能解决问题。图表中列出的模型和方法也能够被应用于研究转向装置的动态伺服系统，可以是系统的设计也可以是动态的分析。 附录 B1 4 New Concept for Hydrostatic Drive with Control of the Secondary Unit Connected to the Ring Main System. Introduction: This thesis proposes a new concept for Control of the Secondary Units Connected to the Hydraulic Ring Main System. In 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 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 motors are controlled by hydraulic means for cost consideration, and energy saving. A major part of this work is directed 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. Background information: 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 into 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. In drive technology, two parameters are important to the power being transmitted: 1) Torque = M (Nm) 2) Speed = n (RPM) These mechanical parameters correspond respectively to the following parameters in hydrostatic drives: 1) Pressure = P (bar) 2) Flow rate = Q (m3/sec) In normal power sources e.g. a combustion engine or an electrical motor, the relation between the mechanical parameters is: P M (1.1) Where: P = Power (KW) 5 M = Torque (Nm) = Angular velocity (rad/sec) For the same 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. Then equation (1.1) becomes: Pmax M K (1.2) Where: K = constant M K/ M1/The relation between torque and speed is vice versa, that is, for any torque there is only one corresponding 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 maximum power produced by the power source, after efficiency considerations. In hydrostatic drives, the story is different. The equation (1.1) becomes: dP Qp (1.3) Where: dp = pressure difference across the motor The hydrostatic transmission is capable of maintaining a preset highpressure level in the hydraulic circuit,while changing the flow rate, by using a pressure regulated pump and a variable displacement motor. Then equation (1.3) becomes: P Q dp QK Where: K = constant P ； Q (1.4) 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 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 motor (dp), 6 remains constant at the maximum value. The characteristic of hydrostatic transmission, is that it can transmit the same maximum torque at any speed, in terms of pressure difference across the motor (dp), by drawing part but not all of the power sources power, which is enough for the preset maximum torque (dp), and producing the required variable speed (Q). Dependent upon the couplings type of the hydraulic parameters two drive concepts can be defined: 1) Drive systems with flow coupling (conventional systems). 2) Drive systems with coupling via the operating pressure (systems with control of the secondary unit). Drive Systems With Flow Coupling (Conventional Systems): The simple hydrostatic drives (one output), consist of the primary unit (pump) and the secondary unit (motor), which are interrelated by virtue of the hydraulic flow Q in (m3/sec). Fig. (1.1) shows this relationship using a closed circuit drive. The volumetric flow Q(which is determined by the input drive speed n1 in rpm and the pump displacement v1) causes the hydraulic motor to achieve an output speed n2 dependent upon its displacement v2. This system is relatively efficient because no throttling elements are present. In heavy engineering applications for multioutput circuits, it is common practice to install hydraulic systems with a central oil supply on a socalled 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 saves weight and cost. The ring main system operates at a constant pressure by using pressureregulated pumps and with many actuators connected in parallel. In order to ensure that all of the fluid does not flow through the actuator with the lowest level of resistance, it is necessary to install throttling elements in the energy transmission lines. These ensure that the relevant amount of flow reaches the individual actuators. There are the ring main system with two actuators in an 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 control 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 while lowering a load, energy is converted into heat in a counter balance valve. Through the throttling and under partially loaded conditions, a part of the power is converted into heat 7 losses. This is in fact a part of the pressure generated by the pumps and is not required by the actuator at any flow. With any change in output torque there is change in pressure drop dp across the actuator； this c hange in pressure drop causes the fluid to compress and then expand again. This has an adverse effect on the system stability due to the HYDRAULIC SPRING. Therefore it is necessary to increase the control time of the pumps to damp the pressure buildup and keep the systems stability under control. Drive Systems With Pressure Coupling (Systems With Control Of The Secondary Unit): It was necessary to look for another 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. The advantages of the secondary control systems are: 1) Parallel operation of a number of actuators without limitation. 2) Energy transfer from the primary to the secondary units without throttling. 3) Energy recovery for use by other actuators or by returning the energy to the primary unit, againwithout throttling. 4) A constant operating pressure in order to eliminate the influence of the hydraulic spring. 5) The ability to include accumulators at any required point. The Drive System For Secondary Control (The New Concept): There are the hydraulic circuit for the secondary control unit connected to the ring main system and fitted with a speed control device. The hydraulic circuit consists of: 1. Variable displacement hydraulic motor. 2. Fixed displacement gear pump. When the hydraulic motor starts running under the reset speed pressure (demand pressure), the swashplate 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 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 swashplate actuator spring force against the preset pressure. When the speed control pump outlet pressure and swashplate actuator spring force exceed the reset speed pressure (demand pressure), the swashplate actuator will start to move in a direction to reduce the hydraulic motor displacement, 8 to reduce the hydraulic motor speed. As the hydraulic motor speed decreases the outlet speed control pump pressure decreases, until the reset speed pressure exceeds the sum of the speed control pump pressure and the swashplate actuator spring force. Then the swashplate actuator starts to move in a direction to increase the hydraulic motor displacement to increase the speed. This action will repeat with decay until the swashplate actuator stops at the equilibrium position, where the summing forces acting on the swashplate actuator equal zero, and the hydraulic motor produces enough torque to maintain the speed, which corresponds to the demand pressure (reset speed pressure). CONCLUSIONS. The simulation for the new concept secondary controlled hydraulic 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 the secondary control the following advantages: 1. Connect many hydraulic drives to the ringmain system, all the hydraulic drives work at individual variable speeds. 2. Energy saving. 3. Low coast. 4. No speed variation at the steady state speeds. 5. Hydraulic controlled drive. 9 References 1. Christophersen, T.O. The Application Of Secondary Control For A Hydraulically Driven Cargo Pump. Thesis for the MSc degree, University of Bath, 1997. 2. Drexler, P. and Faatz, H. Planning And Design Of Hydraulic Power Systems, Mannesmann Rexroth, Gmbh,Lohr am Main, 1988. 3. Feuser, A.and Kordak R. Hydrostatic Drives With Control Of The Secondary Unit, Mannesmann Rexroth, Gmbh, Lohr am Main, 1989. 4. Gawthrop, P. and Smith, L. Met Modelling: Bond graphs and dynamic systems, Prentice Hall International (UK) Limited, 1996. 5. Stecki, J. S. Garbacik, A. and Szewczyk, K. Hydraulic Control Systems, System Design And Analysis, department of mechanical engineering, Monash University, Australia, 1997. 6. Stecki, J. 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. 10 附录 B2 DYNAMIC SIMULATION OF THE HYDRAULIC SERVO SYSTEM FOR AEG STEERING GEAR ABSTRACT A dynamic model of a hydraulic 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 components and matching them by simulation. The findings presented in this paper are valuable for the system designers and analysts. KEY WORDS: Components selection and match； Power bond graphs； The hydraulic servo system for AEG steering gear INTRODUCTION A hydraulic servo system is widely used for steering gears on board ship. It is 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 if 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 examine a hydraulic servo system in detail by simulation. SYSTEM DESCRIPTION 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 through the control valve and the 11 cylinder are shown in Fig.2 Fig. 1 The system Fig. 2 Oil flow in the valve and cylinder DYNAMIC MODELLING (1) Power Bond Graph 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 this 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 rigidly coupled). The capacitance of the line and the pump have been added to the filter, expressed as Cp. Fig. 3 The bond graph SIMULATION In order to solve the 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 has been performed on a computer. Fig.4 shows the input Xv(t). Fig. 5 shows the load displacement Xm(t) 12 Fig. 4 The input displacement of the valve spool Fig. 5 The load displacement Xm(t) COMPARISON AND DISCUSSION 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 specific 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 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 shown 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 the control valve), then most of the liquid will be pumped to the actuator to do work； However, if the pump and the control valve are unmatched, that is, the pump flow rate is greater than the flow rate that the valve permits to flow through, although the control valve stays open, lots of the liquid will have to be pumped to the oil tank through 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. 13 Fig. 6 The flowrate into the control Fig. 7 The flowrate through the relief valve supply port valve to tank Fig.8 The flowrate into the control Fig.9 The flowrate through the relief valve supply port valve to tank CONCLUSIONS The main components of the system should be well chosen especially the pump and the control 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 systems for steering gear both in system design and dynamic analysis. REFERENCES 1 Dang Kun, Power Flow Modeling of the Servo System for AEG Steering Gear, Proceedings of the 3rd Nationwide Chinese Youth Symposium on Mechanical Engineering. Beijing, Mechanical Industry Publisher, 1998,11:135-138 (in Chinese) 2 Dang Kun, Yin Fen, Dynamic Simulation of a Hydraulic Pump Sub-system. Journal of Dalian Maritime University 1998, Vol(24)2:105-107(in Chinese)