公路铣刨机整机的设计开题报告

编号无锡太湖学院毕业设计（论文）相关资料题目： 公路铣刨机整机的设计 信机 系 机械工程及自动化专业学 号： 0923159 学生姓名： 陈双成 指导教师： 何雪明（职称：副教授 ） （职称： ）2013年5月25日目 录一、毕业设计（论文）开题报告二、毕业设计（论文）外文资料翻译及原文三、学生“毕业论文（论文）计划、进度、检查及落实表”四、实习鉴定表无锡太湖学院毕业设计（论文）开题报告题目： 公路铣刨机整机的设计 信机 系 机械工程及自动化 专业学 号： 0923159 学生姓名： 陈双成 指导教师： 何雪明 （职称：副教授 ） （职称： ）2012年11月25日 课题来源本课题来源于工厂。科学依据（1）课题科学意义 沥青混凝土路面铣刨机是一种高效的沥青路面维修养护设备，其原理是利用滚动铣削的方法把沥青混凝土路面局部或全部破碎。铣削下来的沥青碎料经再生处理后，可直接用于路面表层的重新铺筑。主要用于公路、城市道路、机场、货场、停车场等沥青混凝土砼面层开挖翻新；沥青路面拥包、油浪、网纹、车辙等的清除；水泥路面的拉毛及面层错台铣平等。随着市政道路和高等级公路建设突飞猛进，大规模的机械化养护时代已经到来。（2）铣刨机的研究状况及其发展前景 国外路面铣刨机起源于20世界50年代，经过50年的发展，其产品已成系列化，生产效率一般为150-2000，铣刨宽度0.3-4.2m，最大铣刨深度可达350mm，其机电液一体化技术已趋成熟，铣削深度可通过自动找平系统自动控制，同时为改善作业环境，延长铣削刀具的使用寿命，设计有喷洒水装置和密闭转子罩壳。为了减轻劳动强度，近年来开发的产品都带有回收装置，使铣削物从铣削转子直接输送到运载卡车上。国外制造厂商众多，主要有维特根、英格索兰、比泰利、卡特彼勒、戴纳派克等。维特根在国际上处于主导地位，尤其是小型铣刨机更是无人能及。主要生产SF和DC系列铣刨机，已形成了铣刨宽度从0.3-4.2米的近20种规格的产品系列，最大铣削深度为350mm，我国主要以进口该公司产品为主。比泰利已具有40年多制造铣刨机的历史，其SF系列冷铣刨机有11种型号，铣刨宽度为0.6-2.1米，铣刨深度340mm。卡特彼勒主要生产PR和PM两大系列，铣刨宽度为1.9-3.18，铣刨深度305mm，其铣刨机具有铣刨深度和铣刨表面自动调平自动控制功能，铣刨深度误差为3mm。戴纳派克主要生产PL系列铣刨机，铣刨宽度为0.35-2.1米，铣刨深度80-150mm。研究内容 由于国内外已经具有先进的比较完善的铣刨机机型可参考，我们的总体方案设计可以充分利用现有资源，在原有的结构基础上进行类比设计和优化设计。 针对铣刨机的每一个子系统，分析其功能、结构，了解国内外现有的结构， 比较各种机构的优缺点，再结合当前技术的发展，提出新的或改进的系统结构设置。拟采取的研究方法、技术路线、实验方案及可行性分析（1）实验方案 到工厂进行实地观察，仔细了解各部分的结构形式，弄清其工作原理。使用UG画出各个零件，再进行装配、修改，确定正确后，最后进行有限元分析，运动仿真，以检验方案的合理性与可行性。（2）研究方法 实地考查 UG仿真研究计划及预期成果研究计划：2012年11月12日-2012年12月25日：按照任务书要求查阅论文相关参考资料，填写毕业设计开题报告书。2013年1月11日-2013年3月5日：填写毕业实习报告。2013年3月8日-2013年3月14日：按照要求修改毕业设计开题报告。2013年3月15日-2013年3月21日：学习并翻译一篇与毕业设计相关的英文材料。2013年3月22日-2013年4月11日：UG绘图。2013年4月12日-2013年4月25日：仿真，出工程图。2013年4月26日-2013年5月25日：毕业论文撰写和修改工作。预期成果：了解了公路铣刨机的工作原理，基本组成部分，强化了使用UG画图的能力，检验了四年学习的知识，提高了实践能力。特色或创新之处 使用UG画三维图，出工程图，效果明显，方便改变参量，能够直观判断方案的合理性。 采用固定某些参量、改变某些参量来研究问题的方法，思路清晰，简洁明了，行之有效。已具备的条件和尚需解决的问题 实验方案思路已经非常明确，已经具备使用UG绘图的能力和图像处理方面的知识。 使用UG仿真的能力尚需加强。指导教师意见 指导教师签名：年 月 日教研室（学科组、研究所）意见 教研室主任签名： 年 月 日系意见 主管领导签名： 年 月 日英文原文3.1 One Dimensional Mathematical Model 51 The Conservation of Internal Energy （3.1）where is angle of rotation of the main rotor, h = h() is specific enthalpy, m = m () is mass flow rate p = p(), fluid pressure in the working chamber control volume, Q = Q(), heat transfer between the fluid and the compressor surrounding, V = V () local volume of the compressor working chamber. In the above equation the subscripts in and out denote the fluid inflow and outflow.The fluid total enthalpy inflow consists of the following components: (3.2)where subscripts l, g denote leakage gain suc, suction conditions, and oil denotes oil. The fluid total outflow enthalpy consists of: (3.3)where indices l, l denote leakage loss and dis denotes the discharge conditions with m dis denoting the discharge mass flow rate of the gas contaminated with the oil or other liquid injected. The right hand side of the energy equation consists of the following terms which are model The heat exchange between the fluid and the compressor screw rotors and casing and through them to the surrounding, due to the difference in temperatures of gas and the casing and rotor surfaces is accounted for by the heat transfer coefficient evaluated from the expression Nu = 0.023 Re0.8. For the characteristic length in the Reynolds and Nusselt number the difference between the outer and inner diameters of the main rotor was adopted. This may not be the most appropriate dimension for this purpose, but the characteristic length appears in the expression for the heat transfer coefficient with the exponent of 0.2 and therefore has little influence as long as it remains within the same order of magnitude as other characteristic dimensions of the machine and as long as it characterizes the compressor size. The characteristic velocity for the Re number is computed from the local mass flow and the cross-sectional area. Here the surface over which the heat is exchanged, as well as the wall temperature, depend on the rotation angle of the main rotor. The energy gain due to the gas inflow into the working volume is represented by the product of the mass intake and its averaged enthalpy. As such, the energy inflow varies with the rotational angle. During the suction period, gas enters the working volume bringing the averaged gas enthalpy,52 3 Calculation of Screw Compressor Performance which dominates in the suction chamber. However, during the time when the suction port is closed, a certain amount of the compressed gas leaks into the compressor working chamber through the clearances. The mass of this gas, as well as its enthalpy are determined on the basis of the gas leakage equations. The working volume is filled with gas due to leakage only when the gas pressure in the space around the working volume is higher, otherwise there is no leakage, or it is in the opposite direction, i.e. from the working chamber towards other plenums. The total inflow enthalpy is further corrected by the amount of enthalpy brought into the working chamber by the injected oil. The energy loss due to the gas outflow from the working volume is defined by the product of the mass outflow and its averaged gas enthalpy. During delivery, this is the compressed gas entering the discharge plenum, while, in the case of expansion due to inappropriate discharge pressure, this is the gas which leaks through the clearances from the working volume into the neighbouring space at a lower pressure. If the pressure in the working chamber is lower than that in the discharge chamber and if the discharge port is open, the flow will be in the reverse direction, i.e. from the discharge plenum into the working chamber. The change of mass has a negative signand its assumed enthalpy is equal to the averaged gas enthalpy in the pressure chamber. The thermodynamic work supplied to the gas during the compression process is represented by the term pdV d . This term is evaluated from the local pressure and local volume change rate. The latter is obtained from the relationships defining the screw kinematics which yield the instantaneous working volume and its change with rotation angle. In fact the term dV/d can be identified with the instantaneous interlobe area, corrected for the captured and overlapping areas.If oil or other fluid is injected into the working chamber of the compressor, the oil mass inflow and its enthalpy should be included in the inflow terms. In spite of the fact that the oil mass fraction in the mixture is significant, its effect upon the volume flow rate is only marginal because the oil volume fraction is usually very small. The total fluid mass outflow also includes the injected oil, the greater part of which remains mixed with the working fluid. Heat transfer between the gas and oil droplets is described by a first order differential equation.The Mass Continuity Equation （3.4） The mass inflow rate consists of: （3.5）3.1 One Dimensional Mathematical Model 53 The mass outflow rate consists of: (3.6) Each of the mass flow rate satisfies the continuity equation (3.7)where wm/s denotes fluid velocity, fluid density and A the flow crosssectionarea. The instantaneous density = () is obtained from the instantaneous mass m trapped in the control volume and the size of the corresponding instantaneous volume V , as = m/V .3.1.2 Suction and Discharge Ports The cross-section area A is obtained from the compressor geometry and it may be considered as a periodic function of the angle of rotation . The suction port area is defined by: (3.8)where suc means the starting value of at the moment of the suction port opening, and Asuc, 0 denotes the maximum value of the suction port crosssection area. The reference value of the rotation angle is assumed at the suction port closing so that suction ends at = 0, if not specified differently. The discharge port area is likewise defined by: (3.9)where subscript e denotes the end of discharge, c denotes the end of compression and Adis, 0 stands for the maximum value of the discharge port crosssectional area. Suction and Discharge Port Fluid Velocities (3.10)where is the suction/discharge orifice flow coefficient, while subscripts 1 and 2 denote the conditions downstream and upstream of the considered port. The provision supplied in the computer code will calculate for a reverse flow if h2 h1.54 3 Calculation of Screw Compressor Performance3.1.3 Gas Leakages Leakages in a screw machine amount to a substantial part of the total flow rate and therefore play an important role because they influence the process both by affecting the compressor mass flow rate or compressor delivery, i.e. volumetric efficiency and the thermodynamic efficiency of the compression work. For practical computation of the effects of leakage upon the compressor process, it is convenient to distinguish two types of leakages, according to their direction with regard to the working chamber: gain and loss leakages. The gain leakages come from the discharge plenum and from the neighbouring working chamber which has a higher pressure. The loss leakages leave the chamber towards the suction plenum and to the neighbouring chamber with a lower pressure. Computation of the leakage velocity follows from consideration of the fluid flow through the clearance. The process is essentially adiabatic Fanno-flow. In order to simplify the computation, the flow is is sometimes assumed to be at constant temperature rather than at constant enthalpy. This departure from the prevailing adiabatic conditions has only a marginal influence if the analysis is carried out in differential form, i.e. for the small changes of the rotational angle, as followed in the present model. The present model treats only gas leakage. No attempt is made to account for leakage of a gas-liquid mixture, while the effect of the oil film can be incorporated by an appropriate reduction of the clearance gaps. An idealized clearance gap is assumed to have a rectangular shape and the mass flow of leaking fluid is expressed by the continuity equation: (3.11)where r and w are density and velocity of the leaking gas, Ag = lgg the clearance gap cross-sectional area, lg leakage clearance length, sealing line, g leakage clearance width or gap, = (Re, Ma) the leakage flow discharge coefficient.Four different sealing lines are distinguished in a screw compressor: the leading tip sealing line formed between the main and gate rotor forward tip and casing, the trailing tip sealing line formed between the main and gate reverse tip and casing, the front sealing line between the discharge rotor front and the housing and the interlobe sealing line between the rotors. All sealing lines have clearance gaps which form leakage areas. Additionally, the tip leakage areas are accompanied by blow-hole areas. According to the type and position of leakage clearances, five different leakages can be identified, namely: losses through the trailing tip sealing and front sealing and gains through the leading and front sealing. The fifth, “throughleakage” does not directly affect the process in the working chamber, but it passes through it from the discharge plenum towards the suction port. The leaking gas velocity is derived from the momentum equation, which accounts for the fluid-wall friction:3.1 One Dimensional Mathematical Model 55 (3.12)where f(Re, Ma) is the friction coefficient which is dependent on the Reynolds and Mach numbers, Dg is the effective diameter of the clearance gap, Dg 2g and dx is the length increment. From the continuity equation and assuming that T const to eliminate gas density in terms of pressure, the equation can be integrated in terms of pressure from the high pressure side at position 2 to the low pressure side at position 1 of the gap to yield: (3.13)where = fLg/Dg + characterizes the leakage flow resistance, with Lg clearance length in the leaking flow direction, f friction factor and local resistance coefficient. can be evaluated for each clearance gap as a function of its dimensions and shape and flow characteristics. a is the speed of sound. The full procedure requires the model to include the friction and drag coefficients in terms of Reynolds and Mach numbers for each type of clearance. Likewise, the working fluid friction losses can also be defined in terms of the local friction factor and fluid velocity related to the tip speed, density, and elementary friction area. At present the model employs the value of in terms of a simple function for each particular compressor type and use. It is determined as an input parameter.These equations are incorporated into the model of the compressor and employed to compute the leakage flow rate for each clearance gap at the local rotation angle .3.1.4 Oil or Liquid Injection Injection of oil or other liquids for lubrication, cooling or sealing purposes, modifies the thermodynamic process in a screw compressor substantially. The following paragraph outlines a procedure for accounting for the effects of oil injection. The same procedure can be applied to treat the injection of any other liquid. Special effects, such as gas or its condensate mixing and dissolving in the injected fluid or vice versa should be accounted for separately if they are expected to affect the process. A procedure for incorporating these phenomena into the model will be outlined later. A convenient parameter to define the injected oil mass flow is the oil-to-gas mass ratio, moil/mgas, from which the oil inflow through the open oil port, which is assumed to be uniformly distributed, can be evaluated as (3.14)where the oil-to-gas mass ratio is specified in advance as an input parameter56 3 Calculation of Screw Compressor Performance In addition to lubrication, the major purpose for injecting oil into a compressor is to cool the gas. To enhance the cooling efficiency the oil is atomized into a spray of fine droplets by means of which the contact surface between the gas and the oil is increased. The atomization is performed by using specially designed nozzles or by simple high-pressure injection. The distribution of droplet sizes can be defined in terms of oil-gas mass flow and velocity ratio for a given oil-injection system. Further, the destination of each distinct size of oil droplets can be followed until it hits the rotor or casing wall by solving the dynamic equation for each droplet size in a Lagrangian frame, accounting for inertia gravity, drag, and other forces. The solution of the droplet energy equation in parallel with the momentum equation should yield the amount ofheat exchange with the surrounding gas. In the present model, a simpler procedure is adopted in which the heat exchange with the gas is determined from the differential equation for the instantaneous heat transfer between the surrounding gas and an oil droplet. Assuming that the droplets retain a spherical form, with a prescribed Sauter mean droplet diameter dS, the heat exchange between the droplet and the gas can be expressed in terms of a simple cooling law Qo = hoAo(Tgas Toil), where Ao is the droplet surface, Ao = d2 S , dS is the Sauter mean diameter of the droplet and ho is the heat transfer coefficient on the droplet surface, determined from an empirical expression. The exchanged heat must balance the rate of change of heat taken or given away by the droplet per unit time, Qo = mocoildTo/dt = mocoildTo/d, where coil is the oil specific heat and the subscript o denotes oil droplet. The rate of change of oil droplet temperature can now be expressed as: (3.15) The heat transfer coefficient ho is obtained from: (3.16) Integration of the equation in two time/angle steps yields the new oil droplet temperature at each new time/angle step: (3.17)where To,p is the oil droplet temperature at the previous time step and k is the non-dimensional time constant of the droplet, k = /t = /, with = mocoil/hoAo being the real time constant of the droplet. For the given Sauter mean diameter, dS, the non-dimensional time constant takes the form (3.18) The derived droplet temperature is further assumed to represent the average temperature of the oil, i.e. Toil To, which is further used to compute the enthalpy of the gas-oil mixture.3.1 One Dimensional Mathematical Model 57 The above approach is based on the assumption that the oil-droplet time constant is smaller than the droplet travelling time through the gas before it hits the rotor or casing wall, or reaches the compressor discharge port. This means that