喜歡這套資料就充值下載吧。。。資源目錄里展示的都可在線預(yù)覽哦。。。下載后都有，，請放心下載，，文件全都包含在內(nèi)，，【有疑問咨詢QQ:414951605 或 1304139763】
=============================================
喜歡這套資料就充值下載吧。。。資源目錄里展示的都可在線預(yù)覽哦。。。下載后都有，，請放心下載，，文件全都包含在內(nèi)，，【有疑問咨詢QQ:414951605 或 1304139763】
=============================================
編號
無錫太湖學(xué)院
畢業(yè)設(shè)計（論文）
相關(guān)資料
題目： 公路銑刨機整機的設(shè)計
信機 系 機械工程及自動化專業(yè)
學(xué) 號： 0923159
學(xué)生姓名： 陳雙成
指導(dǎo)教師： 何雪明（職稱：副教授 ）
（職稱： ）
2013年5月25日
目 錄
一、畢業(yè)設(shè)計（論文）開題報告
二、畢業(yè)設(shè)計（論文）外文資料翻譯及原文
三、學(xué)生“畢業(yè)論文（論文）計劃、進度、檢查及落實表”
四、實習(xí)鑒定表
無錫太湖學(xué)院
畢業(yè)設(shè)計（論文）
開題報告
題目： 公路銑刨機整機的設(shè)計
信機 系 機械工程及自動化 專業(yè)
學(xué) 號： 0923159
學(xué)生姓名： 陳雙成
指導(dǎo)教師： 何雪明 （職稱：副教授 ）
（職稱： ）
2012年11月25日
課題來源
本課題來源于工廠。
科學(xué)依據(jù)
（1）課題科學(xué)意義
瀝青混凝土路面銑刨機是一種高效的瀝青路面維修養(yǎng)護設(shè)備，其原理是利用滾動銑削的方法把瀝青混凝土路面局部或全部破碎。銑削下來的瀝青碎料經(jīng)再生處理后，可直接用于路面表層的重新鋪筑。主要用于公路、城市道路、機場、貨場、停車場等瀝青混凝土砼面層開挖翻新；瀝青路面擁包、油浪、網(wǎng)紋、車轍等的清除；水泥路面的拉毛及面層錯臺銑平等。隨著市政道路和高等級公路建設(shè)突飛猛進，大規(guī)模的機械化養(yǎng)護時代已經(jīng)到來。
（2）銑刨機的研究狀況及其發(fā)展前景
國外路面銑刨機起源于20世界50年代，經(jīng)過50年的發(fā)展，其產(chǎn)品已成系列化，生產(chǎn)效率一般為150-2000，銑刨寬度0.3-4.2m，最大銑刨深度可達350mm，其機電液一體化技術(shù)已趨成熟，銑削深度可通過自動找平系統(tǒng)自動控制，同時為改善作業(yè)環(huán)境，延長銑削刀具的使用壽命，設(shè)計有噴灑水裝置和密閉轉(zhuǎn)子罩殼。為了減輕勞動強度，近年來開發(fā)的產(chǎn)品都帶有回收裝置，使銑削物從銑削轉(zhuǎn)子直接輸送到運載卡車上。國外制造廠商眾多，主要有維特根、英格索蘭、比泰利、卡特彼勒、戴納派克等。
維特根在國際上處于主導(dǎo)地位，尤其是小型銑刨機更是無人能及。主要生產(chǎn)SF和DC系列銑刨機，已形成了銑刨寬度從0.3-4.2米的近20種規(guī)格的產(chǎn)品系列，最大銑削深度為350mm，我國主要以進口該公司產(chǎn)品為主。比泰利已具有40年多制造銑刨機的歷史，其SF系列冷銑刨機有11種型號，銑刨寬度為0.6-2.1米，銑刨深度340mm?？ㄌ乇死罩饕a(chǎn)PR和PM兩大系列，銑刨寬度為1.9-3.18，銑刨深度305mm，其銑刨機具有銑刨深度和銑刨表面自動調(diào)平自動控制功能，銑刨深度誤差為±3mm。戴納派克主要生產(chǎn)PL系列銑刨機，銑刨寬度為0.35-2.1米，銑刨深度80-150mm。
研究內(nèi)容
由于國內(nèi)外已經(jīng)具有先進的比較完善的銑刨機機型可參考，我們的總體方案設(shè)計可以充分利用現(xiàn)有資源，在原有的結(jié)構(gòu)基礎(chǔ)上進行類比設(shè)計和優(yōu)化設(shè)計。
針對銑刨機的每一個子系統(tǒng)，分析其功能、結(jié)構(gòu)，了解國內(nèi)外現(xiàn)有的結(jié)構(gòu)， 比較各種機構(gòu)的優(yōu)缺點，再結(jié)合當(dāng)前技術(shù)的發(fā)展，提出新的或改進的系統(tǒng)結(jié)構(gòu)設(shè)置。
擬采取的研究方法、技術(shù)路線、實驗方案及可行性分析
（1）實驗方案
到工廠進行實地觀察，仔細了解各部分的結(jié)構(gòu)形式，弄清其工作原理。使用UG畫出各個零件，再進行裝配、修改，確定正確后，最后進行有限元分析，運動仿真，以檢驗方案的合理性與可行性。
（2）研究方法
① 實地考查
② UG仿真
研究計劃及預(yù)期成果
研究計劃：
2012年11月12日-2012年12月25日：按照任務(wù)書要求查閱論文相關(guān)參考資料，填寫畢業(yè)設(shè)計開題報告書。
2013年1月11日-2013年3月5日：填寫畢業(yè)實習(xí)報告。
2013年3月8日-2013年3月14日：按照要求修改畢業(yè)設(shè)計開題報告。
2013年3月15日-2013年3月21日：學(xué)習(xí)并翻譯一篇與畢業(yè)設(shè)計相關(guān)的英文材料。
2013年3月22日-2013年4月11日：UG繪圖。
2013年4月12日-2013年4月25日：仿真，出工程圖。
2013年4月26日-2013年5月25日：畢業(yè)論文撰寫和修改工作。
預(yù)期成果：
了解了公路銑刨機的工作原理，基本組成部分，強化了使用UG畫圖的能力，檢驗了四年學(xué)習(xí)的知識，提高了實踐能力。
特色或創(chuàng)新之處
① 使用UG畫三維圖，出工程圖，效果明顯，方便改變參量，能夠直觀判斷方案的合理性。
② 采用固定某些參量、改變某些參量來研究問題的方法，思路清晰，簡潔明了，行之有效。
已具備的條件和尚需解決的問題
① 實驗方案思路已經(jīng)非常明確，已經(jīng)具備使用UG繪圖的能力和圖像處理方面的知識。
② 使用UG仿真的能力尚需加強。
指導(dǎo)教師意見
指導(dǎo)教師簽名：
年 月 日
教研室（學(xué)科組、研究所）意見
教研室主任簽名：
年 月 日
系意見
主管領(lǐng)導(dǎo)簽名：
年 月 日
英文原文
3.1 One Dimensional Mathematical Model 51
The Conservation of Internal Energy
（3.1）
where θ is angle of rotation of the main rotor, h = h(θ) is specific enthalpy, m˙ = m˙ (θ) is mass flow rate p = p(θ), fluid pressure in the working chamber control volume, ˙Q = ˙Q(θ), heat transfer between the fluid and the compressor surrounding, ˙V = ˙V (θ) 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 θ 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 sign
and 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 dθ . 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 w[m/s] denotes fluid velocity, ρ – fluid density and A – the flow crosssection
area. The instantaneous density ρ = ρ(θ) is obtained from the instantaneous mass m trapped in the control volume and the size of the corresponding instantaneous volume V , as ρ = 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 θ. The suction port area is defined by:
(3.8)
where suc means the starting value of θ 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 θ is assumed at the suction port closing so that suction ends at θ = 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 μ 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 Performance
3.1.3 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 = lgδg the clearance gap cross-sectional area, lg leakage clearance length, sealing line, δg leakage clearance width or gap, μ = μ(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 ≈ 2δg 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 ζ = fLg/Dg + Σξ characterizes the leakage flow resistance, with Lg clearance length in the leaking flow direction, f friction factor and ξ local resistance coefficient. ζ 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 ζ 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 θ.
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 parameter
56 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 of
heat 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 π, 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 = mocoilωdTo/dθ, 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 = τ/Δt = ωτ/Δθ, with τ = 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 τ 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 heat exchange is completed within the droplet travelling time through the gas during compression. This prerequisite is fulfilled by atomization of the injected oil. This produces sufficiently small droplet sizes to gives a small droplet time constant by choosing an adequate nozzle angle, and, to some extent, the initial oil spray velocity. The droplet trajectory computed independently on the basis of the solution of droplet momentum equation for different droplet mean diameters and initial velocities. Indications are that for most screw compressors currently in use, except, perhaps for the smallest ones, with typical tip speeds of between 20 and 50m/s, this condition is well satisfied for oil droplets with diameters below 50 μm. For more details refer to Stosic et al., 1992.
Because the inclusion of a complete model of droplet dynamics would complicate the computer code and the outcome would always be dependant on the design and angle of the oil injection nozzle, the present computation code uses the above described simplified approach. This was found to be fully satisfactory for a range of different compressors. The input parameter is only the mean Sauter diameter of the oil droplets, dS and the oil properties – density, viscosity and specific heat.
3.1.5 Computation of Fluid Properties
In an ideal gas, the internal thermal energy of the gas-oil mixture is given by:
(3.19)
where R is the gas constant and γ is adiabatic exponent
Hence, the pressure or temperature of the fluid in the compressor working chamber can be explicitly calculated by input of the equation for the oil temperature Toil:
(3.20)
If k tends 0, i.e. for high heat transfer coefficients or small oil droplet size, the oil temperature fast approaches the gas temperature.
In the case of a real gas the situation is more complex, be