CY14-1型斜盘式轴向柱塞泵设计【含14张CAD图纸、三维模型图纸、说明书】
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含14张CAD图纸、三维模型图纸、说明书
CY14
型斜盘式
轴向
柱塞
设计
14
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图纸
三维
模型
说明书
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毕业设计（论文）任务书 题目： 斜盘式轴向柱塞泵设计 一、基本任务及要求： 查阅20篇以上参考文献，设计一斜盘式轴向柱塞泵，其主要技术参数：额定压力：31.5MPa；额定转速：1500r/min；公称排量：63mL/r；变量形式：手动。完成斜盘式轴向柱塞泵总装图及主要零件图，并利用三维软件（SOLIDWORKS、 UG或Pr o/E）进行三维建模，指定零件加工仿真及数控编程。 二、进度安排及完成时间：1. 准备阶段 1周 了解设计内容，明确课题任务及要求，搜集有关技术文献资料，自学CAD/CAM软件和相关设计技术。 2. 确定设计方案 2周 完成文献综述和开题报告，提出解决课题问题的初步方案，并对方案优、缺点进行比较，并分析实施可行性，按实际条件确定方案。 3. 实习 1周 4. 具体设计 9周 斜盘式轴向柱塞泵的总体设计，部件装配图及零件图设计，斜盘式轴向柱塞泵的三维建模，指定零件加工仿真及数控编程。 5. 撰写毕业设计说明书 2周 按湖南工程学院毕业设计说明书相关标准要求撰写毕业设计说明书。 6. 毕业答辩 1周 进行毕业答辩准备，完成毕业答辩。 外文原文：The Analysis of Cavitation Problems in the Axial Piston Pumpshu WangEaton Corporation,14615 Lone Oak Road,Eden Prairie, MN 55344This paper discusses and analyzes the control volume of a piston bore constrained by the valve plate in axial piston pumps. The vacuum within the piston bore caused by the rise volume needs to be compensated by the flow; otherwise, the low pressure may cause the cavitations and aerations. In the research, the valve plate geometry can be optimized by some analytical limitations to prevent the piston pressure below the vapor pressure. The limitations provide the design guide of the timings and overlap areas between valve plate ports and barrel kidneys to consider the cavitations and aerations. _DOI: 10.1115/1.4002058_Keywords: cavitation , optimization, valve plate, pressure undershoots1 IntroductionIn hydrostatic machines, cavitations mean that cavities or bubbles form in the hydraulic liquid at the low pressure and collapse at the high pressure region, which causes noise, vibration, and less efficiency. Cavitations are undesirable in the pump since the shock waves formed by collapsed may be strong enough to damage components. The hydraulic fluid will vaporize when its pressure becomes too low or when the temperature is too high. In practice, a number of approaches are mostly used to deal with the problems: (1) raise the liquid level in the tank, (2) pressurize the tank, (3) booster the inlet pressure of the pump,(4) lower the pumping fluid temperature, and (5) design deliberately the pump itself.Many research efforts have been made on cavitation phenomena in hydraulic machine designs. The cavitation is classified into two types in piston pumps: trapping phenomenon related one (which can be prevented by the proper design of the valve plate) and the one observed on the layers after the contraction or enlargement of flow passages (caused by rotating group designs) in Ref. (1). The relationship between the cavitation and the measured cylinder pressure is addressed in this study. Edge and Darling (2) reported an experimental study of the cylinder pressure within an axial piston pump. The inclusion of fluid momentum effects and cavitations within the cylinder bore are predicted at both high speed and high load conditions. Another study in Ref. (3) provides an overview of hydraulic fluid impacting on the inlet condition and cavitation potential. It indicates that physical properties (such as vapor pressure, viscosity, density, and bulk modulus) are vital to properly evaluate the effects on lubrication and cavitation. A homogeneous cavitation model based on the thermodynamic properties of the liquid and steam is used to understand the basic physical phenomena of mass flow reduction and wave motion influences in the hydraulic tools and injection systems (4). Dular et al. (5, 6) developed an expert system for monitoring and control of cavitations in hydraulic machines and investigated the possibility of cavitation erosion by using the computational fluid dynamics (CFD) tools. The erosion effects of cavitations have been measured and validated by a simple single hydrofoil configuration in a cavitation tunnel. It is assumed that the severe erosion is often due to the repeated collapse of the traveling vortex generated by a leading edge cavity in Ref. (7). Then, the cavitation erosion intensity may be scaled by a simple set of flow parameters: the upstream velocity, the Strouhal number, the cavity length, and the pressure. A new cavitation erosion device, called vortex cavitation generator, is introduced to comparatively study various erosion situations (8).More previous research has been concentrated on the valve plate designs, piston, and pump pressure dynamics that can be associated with cavitations in axial piston pumps. The control volume approach and instantaneous flows (leakage) are profoundly studied in Ref. 9. Berta et al. 10 used the finite volume concept to develop a mathematical model in which the effects of port plate relief grooves have been modeled and the gaseous cavitation is considered in a simplified manner. An improved model is proposed in Ref. 11 and validated by experimental results. The model may analyze the cylinder pressure and flow ripples influenced by port plate and relief groove design. Manring compared principal advantages of various valve plate slots (i.e., the slots with constant, linearly varying, and quadratic varying areas) in axial piston pumps 12. Four different numerical models are focused on the characteristics of hydraulic fluid, and cavitations are taken into account in different ways to assist the reduction in flow oscillations 13.The experiences of piston pump developments show that the optimization of the cavitations/aerations shall include the following issues: occurring cavitation and air release, pump acoustics caused by the induced noises, maximal amplitudes of pressure fluctuations, rotational torque progression, etc. However, the aim of this study is to modify the valve plate design to prevent cavitation erosions caused by collapsing steam or air bubbles on the walls of axial pump components. In contrast to literature studies, the research focuses on the development of analytical relationship between the valve plate geometrics and cavitations. The optimization method is applied to analyze the pressure undershoots compared with the saturated vapor pressure within the piston bore.The appropriate design of instantaneous flow areas between the valve plate and barrel kidney can be decided consequently.2 The Axial Piston Pump and Valve PlateThe typical schematic of the design of the axis piston pump is shown in Fig. 1. The shaft offset e is designed in this case to generate stroking containment moments for reducing cost purposes.The variation between the pivot center of the slipper and swash rotating center is shown as a. The swash angle is the variable that determines the amount of fluid pumped per shaft revolution. In Fig. 1, the nth piston-slipper assembly is located at the angle of . The displacement of the nth piston-slipper assembly along the x-axis can be written as xn = R tan（）sin（）+ a sec（） + e tan() (1) where R is the pitch radius of the rotating group.Then, the instantaneous velocity of the nth piston is xn = R sin（）+ R tan（）cos（）+ R sin（） + e (2) where the shaft rotating speed of the pump is=d / dt.The valve plate is the most significant device to constraint flow in piston pumps. The geometry of intake/discharge ports on the valve plate and its instantaneous relative positions with respect to barrel kidneys are usually referred to the valve plate timing. The ports of the valve plate overlap with each barrel kidneys to construct a flow area or passage, which confines the fluid dynamics of the pump. In Fig. 2, the timingangles of the discharge and intake ports on the valve plate are listed as and . The opening angle of the barrel kidney is referred to as . In some designs, there exists a simultaneous overlap between the barrel kidney and intake/discharge slots at the locations of the top dead center (TDC) or bottom dead center (BDC) on the valve plate on which the overlap area appears together referred to as “cross-porting” in the pump design engineering. The cross-porting communicates the discharge and intake ports, which may usually lower the volumetric efficiency. The trapped-volume design is compared with the design of the cross-porting, and it can achieve better efficiency 14. However, the cross-porting is Fig. 1 The typical axis piston pumpcommonly used to benefit the noise issue and pump stability in practice.3 The Control Volume of a Piston BoreIn the piston pump, the fluid within one piston is embraced by the piston bore, cylinder barrel, slipper, valve plate, and swash plate shown in Fig. 3. There exist some types of slip flow by virtue of relative Fig. 2 Timing of the valve platemotions and clearances between thos e components. Within the control volume of each piston bore, the instantaneous mass is calculated as = (3) where and are the instantaneous density and volume such that themass time rate of change can be given asFig. 3 The control volume of the piston bore (4)where d is the varying of the volume.Based on the conservation equation, the mass rate in the control volume is (5)where is the instantaneous flow rate in and out of one piston.From the definition of the bulk modulus, (6)where Pn is the instantaneous pressure within the piston bore. Substituting Eqs. (5) and (6) into Eq. (4) yields (7)where the shaft speed of the pump is .The instantaneous volume of one piston bore can be calculated by using Eq. (1) as = + R tan（）sin（）+ a sec（） + e tan() (8) where is the piston sectional area and is the volume of each piston, which has zero displacement along the x-axis (when =0, ).The volume rate of change can be calculated at the certain swash angle, i.e., =0, such that (9) in which it is noted that the piston bore volume increases or decreases with respect to the rotating angle of .Substituting Eqs. (8) and (9) into Eq. (7) yields(10)4 Optimal DesignsTo find the extrema of pressure overshoots and undershoots in the control volume of piston bores, the optimization method can be used in Eq. (10). In a nonlinear function, reaching global maxima and minima is usually the goal of optimization. If the function is continuous on a closed interval, global maxima and minima exist. Furthermore, the global maximum (or minimum) either must be a local maximum (or minimum) in the interior of the domain or must lie on the boundary of the domain. So, the method of finding a global maximum (or minimum) is to detect all the local maxima (or minima) in the interior, evaluate the maxima (or minima) points on the boundary, and select the biggest (or smallest) one. Local maximum or local minimum can be searched by using the first derivative test that the potential extrema of a function f( ), with derivative , can solve the equation at the critical points of =0 15.The pressure of control volumes in the piston bore may be found as either a minimum or maximum value as dP/ dt=0. Thus, letting the left side of Eq. (10) be equal to zero yields (11)In a piston bore, the quantity of offsets the volume varying and then decreases the overshoots and undershoots of the piston pressure. In this study, the most interesting are undershoots of the pressure, which may fall below the vapor pressure or gas desorption pressure to cause cavitations. The term of in Eq. (11) has the positive value in the range of intake ports (), shown in Fig. 2, which means that the piston volume arises. Therefore, the piston needs the sufficient flow in; otherwise, the pressure may drop.In the piston, the flow of may get through in a few scenarios shown in Fig. 3: (I) the clearance between the valve plate and cylinder barrel, (II) the clearance between the cylinder bore and piston, (III) the clearance between the piston and slipper, (IV) the clearance between the slipper and swash plate, and (V) the overlapping area between the barrel kidney and valve plate ports. As pumps operate stably, the flows in the as laminar flows, which can be calculated as 16 (12)where is the height of the clearance, is the passage length,scenarios IIV mostly have low Reynolds numbers and can be regarded is the width of the clearance (note that in the scenario II, =2 r, in which r is the piston radius), and is the pressuredrop defined in the intake ports as =- (13)where is the case pressure of the pump. The fluid films through the above clearances were extensively investigated in previous research. The effects of the main related dimensions of pump and the operating conditions on the film are numerically clarified inRefs. 17,18. The dynamic behavior of slipper pads and the clearance between the slipper and swash plate can be referred to Refs.19,20. Manring et al. 21,22 investigated the flow rate and load carrying capacity of the slipper bearing in theoretical and experimental methods under different deformation conditions. A simulation tool called CASPAR is used to estimate the nonisothermal gap flow between the cylinder barrel and the valve plate by Huang and Ivantysynova 23. The simulation program also considers the surface deformations to predict gap heights, frictions, etc., between the piston and barrel and between the swash plate and slipper. All these clearance geometrics in Eq. (12) are nonlinear and operation based, which is a complicated issue. In this study, the experimental measurements of the gap flows are preferred. If it is not possible, the worst cases of the geometrics or tolerances with empirical adjustments may be used to consider the cavitation issue, i.e., minimum gap flows.For scenario V, the flow is mostly in high velocity and can be described by using the turbulent orifice equation as (14) where Pi and Pd are the intake and discharge pressure of the pump and and are the instantaneous overlap area between barrel kidneys and inlet/discharge ports of the valve plate individually.The areas are nonlinear functions of the rotating angle, which is defined by the geometrics of the barrel kidney, valve plate ports, silencing grooves, decompression holes, and so forth. Combining Eqs. (11) (14), the area can be obtained as(15)where is the total overlap area of =, and is defined as In the piston bore, the pressure varies from low to high while passing over the intake and discharge ports of the valve plates. It is possible that the instantaneous pressure achieves extremely low values during the intake area( shown in Fig. 2) that may be located below the vapor pressure , i.e., ;then cavitations can happen. To prevent the phenomena, the total overlap area of might be designed to be satisfied with(16)where is the minimum area of = and is a constant that is Vapor pressure is the pressure under which the liquid evaporates into a gaseous form. The vapor pressure of any substance increases nonlinearly with temperature according to the ClausiusClapeyron relation. With the incremental increase in temperature, the vapor pressure becomes sufficient to overcome particle attraction and make the liquid form bubbles inside the substance. For pure components, the vapor pressure can be determined by the temperature using the Antoine equation as , where T is the temperature, and A, B, and C are constants 24.As a piston traverse the intake port, the pressure varies dependent on the cosine function in Eq. (10). It is noted that there are some typical positions of the piston with respect to the intake port, the beginning and ending of overlap, i.e., TDC and BDC ( ) and the zero displacement position ( =0). The two situations will be discussed as follows:(1) When, it is not always necessary to maintain the overlap area of because slip flows may provide filling up for the vacuum. From Eq. (16), letting =0,the timing angles at the TDC and BDC may be designed as (17)in which the open angle of the barrel kidney is . There is no cross-porting flow with the timing in the intake port.(2) When =0, the function of cos has the maximum value, which can provide another limitation of the overlap area to prevent the low pressure undershoots such that (18)where is the minimum overlap area of .To prevent the low piston pressure building bubbles, the vapor pressure is considered as the lower limitation for the pressure settings in Eq. (16). The overall of overlap areas then can be derived to have a design limitation. The limitation is determined by the leakage conditions, vapor pressure, rotating speed, etc. It indicates that the higher the pumping speed, the more severe cavitation may happen, and then the designs need more overlap area to let flow in the piston bore. On the other side, the low vapor pressure of the hydraulic fluid is preferred to reduce the opportunities to reach the cavitation conditions. As a result, only the vapor pressure of the pure fluid is considered in Eqs. (16)(18). In fact, air release starts in the higher pressure than the pure cavitation process mainly in turbulent shear layers, which occur in scenario V. Therefore, the vapor pressure might be adjusted to design the overlap area by Eq. (16) if there exists substantial trapped and dissolved air in the fluid.The laminar leakages through the clearances aforementioned are a tradeoff in the design. It is demonstrated that the more leakage from the pump case to piston may relieve cavitation problems.However, the more leakage may degrade the pump efficiency in the discharge ports. In some design cases, the maximum timing angles can be determined by Eq. (17)to not have both simultaneous overlapping and highly low pressure at the TDC and BDC.While the piston rotates to have the zero displacement, the minimum overlap area can be determined by Eq. 18, which may assist the piston not to have the large pressure undershoots during flow intake.6 ConclusionsThe valve plate design is a critical issue in addressing the cavitation or aeration phenomena in the piston pump. This study uses the control volume method to analyze the flow, pressure, and leakages within one piston bore related to the valve plate timings. If the overlap area developed by barrel kidneys and valve plate ports is not properly designed, no sufficient flow replenishes the rise volume by the rotating movement. Therefore, the piston pressure may drop below the saturated vapor pressure of the liquid and air ingress to form the vapor bubbles. To control the damaging cavitations, the optimization approach is used to detect the lowest pressure constricted by valve plate timings. The analytical limitation of the overlap area needs to be satisfied to remain the pressure to not have large undershoots so that the system can be largely enhanced on cavitation/aeration issues.In this study, the dynamics of the piston control volume is developed by using several assumptions such as constant discharge coefficients and laminar leakages. The discharge coefficient is practically nonlinear based on the geometrics, flow number, etc. Leakage clearances of the control volume may not keep the constant height and width as well in practice due to vibrations and dynamical ripples. All these issues are complicated and very empirical and need further consideration in the future. The results presented in this paper can be more accurate in estimating the cavitations with these extensive studies.Nomenclature the total overlap area between valve plate ports and barrel kidneysAp = piston section area A, B, C= constantsA= offset between the piston-slipper joint and surface of the swash plate = orifice discharge coefficiente= offset between the swash plate pivot and the shaft centerline of the pump = the height of the clearance = the passage length of the clearance M= mass of the fluid within a single piston (kg)N= number of pistonsn = piston and slipper counter = fluid pressure and pressure drop (bar)Pc= the case pressure of the pump (bar)Pd= pump discharge pressure (bar)Pi = pump intake pressure (bar)Pn = fluid pressure within the nth piston bore (bar)Pvp = the vapor pressure of the hydraulic fluid(bar)qn, qLn, qTn = the instantaneous flow rate of each piston(l/min)R = piston pitch radius r = piston radius (mm)t=time (s)V= volumewk = the width of the clearance (mm)x,x= piston displacement and velocity along the shaft axis (m, m/s)=Cartesian coordinates with an origin on the shaft centerline= Cartesian coordinates with an origin on swash plate pivot=swash plate angle and velocity (rad, rad/s)= fluid bulk modulus (bar)= timing angle of valve plates at the BDC and TDC (rad) = the open angle of the barrel kidney(rad)= fluid density(kg/m3) = angular position and velocity of the rotating kit (rad, rad/s) =absolute viscosity(Cp)= coefficients related to the pressure drop外文中文翻译：在轴向柱塞泵气蚀问题的分析本论文讨论和分析了一个柱塞孔与配流盘限制在轴向柱塞泵的控制量设计。真空是由柱塞的运动量引起的，需要由流动补偿，否则，低气压可能导致的气蚀和曝气。配流盘几何的研究，可以优化一些分析性的限制，以防止蒸气压以下的柱塞压力。配流盘的端口和缸体腰形窗口之间重叠的地方，设计时要考虑的空蚀和曝气。关键词：空蚀，优化，配流盘，负脉冲压力1介绍在水压机等液压元件中，空穴或气穴意味着，在低压区液压液体会出有空腔或气泡形成以及崩溃在高压地区，这将导致噪声，振动，这将会降低效率。空蚀对泵的使用是极为不利的，这是因为倒塌形成的冲击波可能像炸弹一样足以损坏元件。当其压力过低或温度过高时，液压油会蒸发。在实践中，许多方法大多用于处理这些问题，比如：（1）提高油箱中的液位高度，（2油箱加压，（3）提高泵的进口压力，（4）降低泵内流体的温度，（5）特意设计的柱塞泵本身，对其结构进行优化设计。在液压机设计中的气蚀现象，许多研究成果已取得一定的成果。在柱塞泵中，气蚀主要可以分为两种类型：一是与困油现象有关（这种现象可通过适当的设计配流盘来阻止困油现象的发生）和所观察到的层上收缩或扩大后的流动通道(由于旋转设计所造成的)。在这项研究中处理气蚀和测量气缸压力之间的关系。Edge and Darling报道了关于轴向柱塞泵内的气缸压力的实验研究。其中包括流体势效应和气蚀在气缸内高速度和高负荷条件的预测。另一项研究概述了液压流体影响进气条件和汽蚀潜力的观点。它表明,物理属性(如蒸汽压力、粘度、密度和体积弹性模量)对适当地评估影响润滑和气蚀是至关重要的。一个相似的气蚀模型在热力学性质液体和蒸汽基础上的用来理解了基本的物理现象的质量流量减少和波动产生影响的液压工具和喷射系统。Dular et al开发了一套专业系统用它来监测和控制的液压机械和调查气蚀的可能性通过使用运用计算流体动力学(CFD)工具。通过一个简单的单翼配置在一个空化隧道，气蚀侵蚀作用已经被测量和验证。Alpha它假定了严重侵蚀经常是由于一个主要的空穴飞转的漩涡重复的崩溃所产生的。然后,在汽蚀强度通过一套简单流参数可能扩大:上游速度 ,空腔长度和压力。一个新的空蚀装置,称为漩涡汽蚀生成器,介绍了各种侵蚀情况。更多的先前的研究已经被集中在阀板的设计,在轴向柱塞泵中活塞和泵压动力学与空穴现象相关联。控制体积的方法和瞬时流(泄漏)正在深刻地研究中。Berta et al采用有限体积的概念发展了一个数学模型，压力平衡槽的形式已经被效仿和气态的汽蚀被认为是在一个简化的方式。一种改进的模型已经被提出且实验验证了其结果。该模型可以分析气缸压力和流量的涟漪影响压力平衡槽的设计。四种不同的数值模型的重点是液压液体的特点,考虑到空穴以不同的方法协助减少流量振荡。柱塞泵发展的经验表明,优化的空穴现象应当包括下列问题:发生气蚀和空气释放、泵声学引起的噪声诱导、最大振幅的压力波动,转动力矩进展等。然而,这项研究的目的是修改配流盘的设计来防止气蚀造成侵蚀蒸汽或空气泡沫崩溃的墙壁上的轴向泵组件。与文学研究相反，这项研究主要集中在配流盘的几何形状和气蚀分析之间的关系的发展。此优化方法应用于分析的压力脉冲与活塞孔内饱和蒸汽压。2轴向柱塞泵,配流盘典型轴向柱塞泵的原理如图1所示。在这种情况下，轴偏移e的设计对降低成本是十分重要的。柱塞泵斜盘倾角度是可变的，决定每转的排量即决定柱塞泵的流量 。如图一所示，第N个柱塞滑靴组件转过的转角为在第n个柱塞滑靴组件沿x轴的位移可以写成xn = R tan（）sin（）+ a sec（） + e tan() (1) 其中R为柱塞滑靴组件的分布圆半径。此刻，在第n个柱塞的瞬时速度是xn = R sin（）+ R tan（）cos（）+ R sin（） + e (2) 其中轴泵的旋转速度=d / dt.配流盘是约束柱塞泵流量的最重要的设备。配流盘吸排油窗口的几何形状以及瞬时相对缸体腰形窗口的位置通常被称为配流盘的时间效应。配流盘开口与缸体底部的腰形窗口的重叠构建流程区域或通道,它限制了柱塞泵的流体动力学，影响着其力学性能。在图2中，在配流
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