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公路铣刨机整机的设计【5张CAD图纸+UG+毕业论文】【答辩通过】

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关键词：: 公路铣刨机整机设计

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摘要

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。作为路面养护和再生设备的主要机种之一的路面铣刨机正越来越引起道路养护专家和施工单位的关注。

路面冷铣刨机是一种高效的路面维修养护设备, 具有使用成本低、铣深范围广及不污染环境等优点。主要用于大面积路面铣刨破碎、常见病害的铣平修整以及路面拉毛作业。路面铣刨机的主要工作装置是铣刨转子，铣刨转子上均匀布置着按左右对称螺旋线排列的铣刀。铣刨转子是以垂直进给和水平进给两种方式进行工作。路面铣刨机以工作速度向前移动，铣刨转子旋转均匀分布于滚筒上并按螺旋线排列的铣刨刀，铣刨刀顺序接触路面，路面材料在铣刨转子刀具的冲击和挤压下破碎为颗粒状；同时呈螺旋状对称排列的刀具将铣下的废料向转子中央聚集，通过抛料板将废料抛到输送机上，并转移到指定位置和运输车辆上。

本设计说明书包括了整机的总体设计和计算，动力机构及行星减速器的设计，行走机构的设计，铣刨部分的设计，输料装置的设计，并对路面铣刨机械的技术发展趋势和市场前景做了简单介绍。

本说明书与图纸配套使用。

关键词：铣刨机；行星减速器；履带；输料装置；刀具

Abstract

With rapid advances in municipal roads and the construction of high-grade highways, mass maintenance time has come. As road maintenance and recycling equipment of main machine kind of pavement milling machine is one of a growing source of concern for experts of road maintenance and construction units.

Cold milling machine is a high performance pavement maintenance equipment, with low cost, wide range of milling depth and does not pollute the environment and so on. Mainly used for common disease of large area of broken pavement milling, milling picking jobs in flat trim as well as pavement. Road milling machine milling rotor are the main working device, evenly milled rotor layout arranged in a symmetrical helical milling cutters. Rotor is based on vertical and horizontal milling feed work in two ways. Pavement milling planing machine to work speed forward mobile, milling planing rotor rotating uniform distribution Yu drum Shang and by spiral arranged of milling Shaver, milling Shaver order contact pavement, pavement material in milling planing rotor tool of impact and extrusion Xia broken for particles shaped; while is spiral shaped symmetric arranged of tool will milling Xia of waste to rotor Central gathered, through throwing material Board will waste throwing to conveying machine Shang, and transfer to specified location and transport vehicles Shang.

The design specifications, including the the overall programme, power and design of the planetary gear reducer, mechanism design, milling parts design, design of the feeding device, and pavement milling machine technology trends and market prospects of doing a simple introduction.

Supporting the use of these instructions with drawings.

Keywords: milling planer ;planetary reducer ;track ;transporting machine ;tool

摘要III

AbstractIV

1 绪论1

1.1路面铣刨机及其发展概况1

1.1.1路面铣刨机的分类1

1.1.2路面铣刨机的应用特点2

1.1.3 路面铣刨机的主要结构及工作原理2

1.1.4 国内外路面铣刨机发展概况3

1.2 路面铣刨机设计的指导思想5

1.3 路面铣刨机设计的设计原则5

1.4本文设计的主要内容5

2 路面铣刨机的总体设计6

2.1 路面铣刨机的选型6

2.2路面铣刨机的总体参数的选型6

2.3 路面铣刨机的总体布置8

2.4 铣刨机总体方案的设计8

3 铣刨机动力机构及行星减速器的设计10

3.1 发动机的选择10

3.2 传动形式的选择11

3.3 行星减速器的设计13

3.3.1 型行星减速器的运动设计13

3.3.2 减速器的结构设计17

4 铣刨机行走系统的设计20

4.1 铣刨机行走系统的基本要求20

4.2 铣刨机行走驱动形式20

4.3 导向与张紧缓冲装置的设计参数分析21

4.4 履带式铣刨机的运动学与动力学分析25

4.4.1 铣刨机行驶系统的运动学分析25

4.4.2 履带式铣刨机动力学分析26

4.5 附着条件决定的最大切线牵引力27

4.6 行走系统功率分析27

5 铣刨机铣刨系统的设计29

5.1 路面铣刨机铣刨转子排列方式分析29

5.2 路面铣刨机铣刨轮刀具排列参数分析30

5.3 路面铣刨机铣刨轮主要参数的研究34

5.3.1 铣刨轮的切削量34

5.3.2 铣刨轮直径35

5.3.3 铣刨轮最佳螺旋升角的选取36

6 输料系统的设计38

6.1 输料系统的主要功能部件38

6.2 输送带的跑偏及防偏措施40

6.2.1 跑偏的危害及原因40

6.2.2 输送带跑偏的常见处理方式40

7 总结与展望43

7.1 总结43

7.2不足之处及未来展望43

致谢45

参考文献46

附录48

1 绪论

1.1 路面铣刨机及其发展概况

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方式把沥青路面局部或全部破碎。铣削下来的沥青碎料经处理后，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。作为路面养护和再生设备的主要机种之一的路面铣刨机正越来越引起道路养护专家和施工单位的关注。公路建设部门对路面铣刨机等成套设备的需求会越来越迫切，需求量也会越来越多。

1.1.1路面铣刨机的分类

按照铣刨宽度，路面铣刨机可分为大中小三种，0.3~1.0米为小型；1.2~2.0米为中型，2.0米以上为大型。对于铣刨宽度小于或等于1.0米的小型铣刨机主要是机械传动，机械式传动工作可靠、维修方便、传动效率高、制造成本低，但其结构复杂、操作不轻便、作业效率较低、牵引力较小，一般适用于小面积的路面维修、刮除喷涂标线、铣刨小型沟槽等，一般不带废料回收装置。铣刨宽度在1.2米以上的中型铣刨机主要是液压传动，液压式结构紧凑、操作轻便、牵引力较大，但制造成本高、维修较难，适用于切削较深的中、大规模路面养护作业。我国目前以生产小型路面铣刨机为主，且以 0.5米和1.0米两种规格居多，大中型产品基本上还是空白。

另外，按铣刨型式不同，可分为热铣和冷铣。热铣式因为增加了加热装置而使结构比较复杂，一般用于路面再生作业。冷铣由于适用范围广，目前占据主导地位。冷铣式配置功率比较大，刀具磨损比较快，切削料粒度较为均匀，可设置洒水装置喷水，使用广泛，且产品成系列；按铣刨机按行走方式不同，可分为轮式和履带式。轮式的机动性好、换场方便，特别适合于中小型的路面作业。履带式多为铣削宽度2000mm以上的大型铣刨机，有旧材料回收装置，适用于大面积路面再生工程；按铣刨鼓旋转方向与行走方向不同，可分为逆铣和顺铣；按有无废料回收输送机，可分为输送式和无输送式；按铣刨鼓安装方式不同，可分为固定式和移动式。

图1.1 轮式铣刨机

图1.2 2m履带式铣刨机

1.1.2 路面铣刨机的应用特点

1. 使用铣刨机铣削路面，可以快速有效地处理路面病害，使路面保持平整；

2. 道路的翻修工程采用铣削工艺可保持原路面的水平高程。铣削工艺可将损坏路面切除掉，由新材料填补原有空间，经压实后与原路面等高，保持路面的原有水平高程，这使穿行于高架桥或立交桥涵的路面载荷对桥体不致产生冲击载荷，并且桥涵通过高程不变；

3. 保证新旧路面材料的良好结合，提高其使用寿命。采用铣削工艺可使填料坑边侧及底部整齐、深度均匀，形成新旧料易于结合的齿状几何表面，从而使翻修后新路面的使用寿命大大提高；

4. 有利于旧路面材料的再生利用。由于可以掌握切削深度，铣削下来的材料不仅干净且呈规则的小颗粒，可以不用再破碎加工即可遮现场或固定料场再生利用，大大降低了施工成本，同时也是一种环境保护措施。

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内容简介：: 存档编码：无无锡锡太太湖湖学学院院 2013 届届毕毕业业作作业业周周次次进进度度计计划划、检检查查落落实实表表系别：信机系班级：机械94 学生姓名：陈双成课题（设计）名称：公路铣刨机整机的设计开始日期：2012年11月12日周次起止日期工作计划、进度每周主要完成内容存在问题、改进方法指导教师意见并签字备注1-32012年11月12日-2012年12月2日教师下达毕业设计任务，学生初步阅读资料，完成毕业设计开题报告。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存在问题：对课题理解程度不够，对其难点分析不够，分析能力欠缺。改进方法：在指导老师的帮助下，对课题有较深的了解。4-102012年12月3日-2013年1月20日指导专业实训机械设计综合实训，机械关键部件设计校核存在问题：机械部件设计不够完善，缺少经验。改进方法：了解机械设计的详细过程。11-122013年1月21日-3月1日指导毕业实习相关机械制造厂实习，了解本专业的实践知识存在问题：没有实习实训的经验，无法将课本知融会贯通。改进方法：认真参与工作，虚心求教。132013年3月4日-3月8日查阅参考资料查阅与设计有关的参考资料不少于10篇，其中外文不少于5篇存在问题：查阅资料相关度小，无法满足要求。改进方法：利用空余时间，去图书馆查找相关资料，在网上查找相关文献。142013年3月11日-3月15日翻译外文资料翻译机械方面的外文资料存在问题：专业英文水平较低，无法正确翻译专业词汇。改进方法：借助一些翻译软件、专业字典帮助翻译提高翻译准确性性。152013年3月18日-3月22日公路铣刨机选型整体设计方案分析各种公路铣刨机的性能，查阅相关资料，根据实际选择存在问题：缺乏设计经验，传动方式不合理。改进方法：重新确立合理的传动方案。162013年3月25日-3月29日绘制装配图初步绘制装配图存在问题：零部件设计有错误。改进方法：重新设计校核并绘制相关部件。172013年4月1日-4月5日绘制装配图初步完成装配图存在问题：不能完全体现铣刨机传动细节。改进方法：绘制局部视图，体现传动过程。182013年4月8日-4月12日主要零件结构设计和校核对主要零件结构和尺寸进行设计和计算存在问题：主体框架结构设计不合理，无法承受载荷。改进方法：认真查阅各方面资料，重新校核计算。周次起止日期工作计划、进度每周主要完成内容存在问题、改进方法指导教师意见并签字备注192013年4月15日-4月20日绘制完整装配图修改完成装配图存在问题：不能完全体现铣刨机各部组成细节。改进方法：绘制局部视图，体现各部细节。202013年4月22日-4月27日绘制零件图绘制整体机架确定定位尺寸存在问题：标注尺寸不全，无图纸标号。改进方法：检查标注，补全图号。212013年4月29日-5月3日绘制零件图绘制车辆底盘存在问题：标注尺寸不全，图纸表示不全面。改进方法：检查标注，绘制局部剖图。222013年5月6日-5月10日设计说明书（论文）、摘要和小结编写完成设计说明书（论文）、摘要和小结存在问题：说明书的格式不规范，摘要不合理要求等。改进方法：根据毕业设计的规范要求更改，重新按要求编写摘要。23-252013年5月13日-5月25日修改设计说明书（论文）格式上交资料、准备答辩修改设计说明书开题报告格式存在问题：附录格式不规范，摘要英文不合理要求等。改进方法：根据毕业设计的规范要求更改。说明： 1、“工作计划、进度”、“指导教师意见并签字”由指导教师填写，“每周主要完成内容”，“存在问题、改进方法”由学生填写。 2、本表由各系妥善归档，保存备查。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 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 m. 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 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, because the temperature and pressure can not be calculated explicitly. However, since the internal energy can be expressed as a function of the temperature and specific volume only, the calculation procedure can be simplified by employing the internal energy as a dependent variable instead of enthalpy, as often is the practice. The equation of state p = f1(T,V ) and the equation for specific internal energy u = f2(T,V ) are usually decoupled. Hence, the temperature can be calculated from the known specific internal energy and the specific volume obtained from the solution of differential equations, whereas the pressure中文译文33.1一维数学模型 51内部能量守恒 (3.1)其中是角度的旋转的主旋翼h =h（）的比焓，m =m （）是质量流率p = （），工作腔的控制体积中的流体压力， Q = Q（）的流体之间的热传递和压缩机周围， V = V （），压缩机工作腔中的本地卷。在上述方程中，输入和输出的下标表示的流体流入及流出。流体的总焓流入由以下组件： (3.2)其中，下标L，G表示泄漏增益SUC ，抽吸条件，和油为石油。流体总流出焓包括： (3.3)指数升， l表示泄漏损耗和dis表示放电条件与m显示表示放电注入的油或其它液体污染的气体的质量流率右手法侧的能量方程由模型的下列术语流体和压缩机的螺杆转子和壳体，并通过它们的周边，由于气体的温度的差异，上述壳体和转子的表面之间的热交换的传热系数求值表达式= 0.023， RE0占.8 。通过主转子的外径和内径之间的差异为特征长度的雷诺数和努塞尔数。这可能不是用于此目的的最合适的尺寸，但出现的特征长度在0.2的指数部分的传热系数的表达式，因此，只要它表征压缩机的体积，它仍然在同一个数量级，作为其他特征尺寸的影响不大的机器。特征速度为Re数的计算从本机的质量流量和横截面面积。这里的表面，在其上进行热交换，以及壁温，依靠的主旋翼的旋转角度。上述所表示的商品的大量摄入量和其平均焓由于工作体积的气体流入的能量增益决定。因此，能量的流入的旋转角变化。在吸入期间，等于气体进入工作容积带来的平均气体焓。52 3螺杆压缩机性能的计算吸入室中占主导地位。然而，在吸入口关闭时，一定量的压缩气体通过间隙泄漏到压缩机工作腔。该气体的质量，以及其焓在气体泄漏方程的基础上确定。工作体积充满了气体，由于泄漏，只有当工作体积周围的空间中的气体压力较高，否则无泄漏，或它是在相反的方向，即从对其他压力通风系统的工作腔。总流入焓进一步校正的焓的量带入工作腔注入的油。由于从工作体积的气体流出的能量损失是指由商品质量的流出和平均气体焓。在工作过程中，这是进入排放气室，被压缩的气体的同时，在扩展的情况下，由于不适当的排出压力，这是通过在较低压力下工作体积到邻近的空间的间隙泄漏的气体。如果工作腔中的压力低于在排出室，排放口是打开的，该流程将在相反的方向，即从排出气室进入工作腔。质量的变化，有一个负号其假定的焓等于压力腔中的平均气体焓。供给的工作气体在压缩过程中的热力学表示由术语PdV d 。这个术语是从本地的压力和体积变化率进行评估。后者被定义产生瞬时工作体积和其旋转角度的变化的螺杆运动学的关系得到的。事实上，术语的dV /差d可确定瞬时interlobe区，捕获和重叠区域校正。如果油或其它流体注入上述压缩机的工作腔，油质量的流入和其焓应包括在流入条款而事实，尽管在混合物中的油的质量分数显着的体积流率时，其效果是不明显的，因为油的体积分数通常是非常小的。总流出的流体的质量，还包括注入的油，其中的较大部分仍然与工作流体混合。气体之间的热传递和油滴描述由一个一阶微分方程确定。质量连续性方程 (3.4)质量连续性方程 (3.5)3.1一维数学模型 53质量的流出率包括： (3.6)质量流率的每一个方程满足连续性方程 (3.7)其中W m/s表示流体速度， - 流体密度和A - 流体截面区域。得到的瞬时密度 = （）被困在控制量与相应的瞬时体积V的大小从瞬时的质量为m ，密度为 =m/ V 。3.1.2吸气和排气口从压缩机的几何形状的横截面面积A得到的旋转角度，它可以被认为是周期函数。吸气口区域被定义为： (3.8)SUC装置上面的吸气口开口，并且ASUC的时刻开始的值，0表示为在吸入口的横截面面积的最大值。如果未指定不同的旋转角度的基准值，吸入口关闭时，假设在吸管末端 = 0。排放口区同样被定义为：（3.9）其中下标e表示放电结束， c表示排出口的横截面面积的最大值压缩和ADIS ， 0表示结束。吸入和排出端口流体速度（3.10）其中，为吸入/排放孔的流量系数，而下标1和2表示所考虑的端口的上游和下游，在计算机代码中提供计算，如果H2 H1反向流动。54 3螺杆压缩机性能的计算3.1.3气体泄漏泄漏量的主要部分是总流速的螺纹机，因此发挥了重要作用，因为它们影响的过程都影响了压缩机的质量流率或压缩机送货，即容积效率和压缩工作的热力学效率。对于实际计算时压缩机的过程中泄漏的影响，这是方便区分的两种类型泄漏，根据他们的方向方面的工作室：增益和损失的泄漏。增益来自排放气室，并从相邻的工作腔室获得，其中有一个较高的压力泄漏。亏损泄漏离开吸气室和邻近腔室向具有较低的压力的腔室流动。泄漏速度的计算如下考虑的流体流过的间隙。该过程本质上是绝热的Fanno流。为了简化计算，该流程是有时被假设为在恒定的温度条件下，而不是在等焓。此处出发从当时的绝热条件下进行分析以差的形式，小的旋转角的变化来表示，即在本模型中，泄漏只有很轻微的影响。本模型只考虑气体泄漏，没有尝试考虑到泄漏的气 - 液混合物中，可掺入适当减少间隙的间隙油膜的影响效果。一个理想化的间隙被假定为

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