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毕业设计（论文）任务书I、毕业设计(论文)题目：液压挖掘机反铲工作装置设计II、毕 业设计(论文)使用的原始资料(数据)及设计技术要求：主要对由动臂、斗杆、铲斗、销轴、连杆机构组成挖掘机工作装置进行设计。具体内容包括以下五部分:(1) 挖掘机工作装置的总体设计。 (2) 挖掘机的工作装置详细的机构运动学分析，运动模拟。(3) 工作装置各部分的基本尺寸的计算和验证。(4) 工作装置主要部件的结构设计。仿真设计。(5) 销轴的设计及螺栓等标准件进行选型。III、毕 业设计(论文)工作内容及完成时间：1. 查找资料，外文资料翻译（不少于6000字符），开题报告 第1周-第2周2运动及动力参数计算 第3周-第4周3总装配图设计 第5周-第6周 4. 工作装置各部分基本尺寸设计 第7周-第8周5用UG/Solidworks对系统进行实体建模和设计 第9周-第11周6. 绘制零、部件图 第12周-第13周7. 毕业论文撰写 第14周-第16周8 .答辩准备及论文答辩 第17周-第17周 、主 要参考资料：1成大先主编.机械设计手册(第三版,第三卷),第十四篇M.化学工业出版社,1992年2付 越，邓子龙. 基于ProE的液压挖掘机反铲工作装置运动仿真，辽宁石油化工大学学报.2007.6.pp50-53.3 邓子龙，刘杰，高财禄等.挖掘机铲斗结构优化.机械与电子.2009.1.pp13-16.4 李滨城，何允纪.液压挖掘机反铲工作装置运动的模拟.华东船舶工业学院学报.1995.6.pp74-805 刘玉强，王学军.液压挖掘机反铲工作装置优化设计，机械产品与科技.1997.1.pp12-156 Yu. I. Berezhnoi and Yu. A. Potapov. Method for the intermittent optimization of the working of a block by a rotary excavator Refractories and Industrial Ceramics, 1982, Volume 23, Numbers 3-4, Pages 183-186 航空与机械工程 系 机械设计制造及其自动化 专业类 班 学生（签名）： 填写日期： 2011 年 01 月 03 日指导教师（签名）： 助理指导教师(并指出所负责的部分)：机械设计制造及其自动化 系主任（签名）：学士学位论文原创性声明本人声明，所呈交的论文是本人在导师的指导下独立完成的研究成果。除了文中特别加以标注引用的内容外，本论文不包含法律意义上已属于他人的任何形式的研究成果,也不包含本人已用于其他学位申请的论文或成果。对本文的研究作出重要贡献的个人和集体，均已在文中以明确方式表明。本人完全意识到本声明的法律后果由本人承担。作者签名： 日期：学位论文版权使用授权书本学位论文作者完全了解学校有关保留、使用学位论文的规定，同意学校保留并向国家有关部门或机构送交论文的复印件和电子版，允许论文被查阅和借阅。本人授权南昌航空大学科技学院可以将本论文的全部或部分内容编入有关数据库进行检索，可以采用影印、缩印或扫描等复制手段保存和汇编本学位论文。 作者签名： 日期： 导师签名： 日期：液压反铲装置设计学生姓名：卢越 班级：078105222指导老师：邢普摘 要：本次设计的题目是液压挖掘机反铲装置机构。与其它类型的挖掘机相比，这种类型的挖掘机因有良好通过性能应用最广，对松软地面或沼泽地带还可采用加宽、加长以及浮式履带来降低接地比压。液压挖掘机反铲装置的主要特点为：反铲是中小型液压挖掘机的主要工作装置。液压挖掘机的反铲装置由动臂，斗杆铲斗，以及动臂油缸，斗杆油缸，铲斗油缸和连杆机构组成。其构造特点是各部件之间的连接全部采用铰接，通过油缸的伸缩来实现挖掘工作中的各种动作。动臂的小铰点与回转平台铰接，并以动臂油缸来支撑和改变动臂的倾角，通过动臂油缸的伸缩可使动臂绕小铰点转动而升降。斗杆铰接于动臂的上端，斗杆与动臂的相对位置由斗杆油缸来控制，当斗杆油缸伸缩时，斗杆便可绕动臂上焦铰点转动。铲斗与斗杆前端铰接，并通过铲斗油缸伸缩使铲斗绕该点转动。为增大铲斗的转角，通常以连杆机构与铲斗连接。本次设计的主要参数是斗容量0.2m3,它属于中小型液压挖掘机，主要设计挖掘机的工作装置。在设计中，采用了轮胎式行走装置，来满足要求。上部转台是全回转式，因此它可在一个更大的范围内工作。又因采用液压传动控制而使整机性能得以改善。与机械式挖掘机相比，其挖掘力提高到23倍，整机质量约为5吨，挖掘力约为30kN，最大卸载高度约为2.65m，最大挖掘深度4.2m，最大挖掘半径约为5.728m，从中可以看出整机作业能力有了很大的改进，不仅挖掘力大，且机器重量轻，传动平稳，作业效率高，结构紧凑。另外，还对挖掘机的工作装置提出基于结构推理的机构方案创新设计方法。关键词：液压挖掘机 ；挖掘机构 ；创新设计 指导教师签名：The design of hydraulic excavator shovel device Student name: luyue class:0781052 Supervisor:xingpuABSTRACT:This designed topic is the marching hydraulic excavator excavational organization. Compared with other types excavators, this kind of type excavator used very universal that because has good through theperformance, also may use to lengthens widens as well as the floating type caterpillar band to reduce pressure for the soft ground or the bogregion.The hydraulic excavator main characteristic is: The small and medium-sized hydraulic excavator shovel is the main work device. Hydraulic excavator shovel device by the arm, dou stem bucket, and arm oil cylinder, dou rod oil cylinders, the bucket of cylinder and linkage mechanism. Its structure feature is between components All adopt the connection by oil cylinder hinged adjustable to realize the various movements excavation. Moving arm little hinge point and rotary platform, and with hinged arm oil cylinder to support and change the dip Angle, through arm arm Oil cylinder telescopic can make moving arm around small hinge point lifting rotation. Dou lever arm hinged on the upper arm, dou rod and the relative By dou pole position to control oil cylinder, when dou rod oil cylinder telescopic, dou lever arm can be around the upper energizer hinge point rotation. Bucket and measures Rod front-end loader, and through hinged cylinder telescopic made the bucket turning around the point. To increase the bucket corner, usually by connecting rod Institutions and bucket connection. theperformance good,and may make the high speed reverse, the transmission steady,structure simple, may absorb attacks and vibrates, the operation reduces effort, and to be easy to realize the automated control, is easy to realize the standardization, the seriation, the universalization.This designed main parameter is scoop capacity 0.2m3, it is long to the middle and small scale hydraulic excavator, mainly design the excavator，s the work installment and the hydraulic transmissionprinciple.In the design, used marching walked the installment to satisfied request. Upside the turnplate is the entire rotation , thereof it may work in a greater scope. And further because uses the hydraulicsteering to enable the entire machine performance to improve. Compared with the mechanical type excavator, its excavation strength enhance to 2 3 times, the entire machine weight approximately is 5 tons,the excavation strength approximately is 30kN, the biggest unloading high approximately is 2.65m, biggest digging depth is 4.2m, the biggest excavation radius approximately is 5.728m, thus can see the entire machine work ability to have the very big improvement, not only excavation strength big, but also machine weight light, transmission steadyly, work efficiency is high, the structure is compact. Moreover, but also proposes to the excavator work installment based on the structureinference organization plan innovation design method.Key word：Hydraulic pressure excavator；Excavation organization；Hydraulic system；Innovation design Signature of supervisor:毕业设计（论文）开题报告题目 液压挖掘机反铲工作装置设计专 业 名 称 机械设计制造及其自动化班 级 学 号 078105222学 生 姓 名 卢越 指 导 教 师 邢普填 表 日 期 2011 年 5 月 20日一、选题的依据及意义：液压挖掘机反铲挖掘装置大大改善了挖掘机的技术性能，挖掘力大、牵引力大，机器重量，传动平稳，作用效率高，结构紧凑。液压挖掘机反铲工作装置主要用于挖掘停机面以下的土壤，如挖掘沟壕、基坑等，其挖掘轨迹取决于各液压缸的运动及其组合。反铲液压挖掘机的工作过程为，先下放动臂至挖掘位置，然后转动斗杆及铲斗，当挖掘至装满铲斗时，提升动臂使铲斗离开土壤，边提升边回转至卸载位置，转斗卸出土壤，然后再回转至工作装置开始下一次作业循环。动臂液压缸主要用于调整工作装置的挖掘位置，一般不单独直接挖掘土壤；斗杆挖掘可获得较大的挖掘行程，但挖掘力小一些。转斗挖掘的行程较短，为使铲斗在转斗挖掘结束时装满铲斗，需要较大的挖掘力以保证能挖掘较大厚度的土壤，因此挖掘机的最大挖掘力一般由转斗液压缸实现的。由于挖掘力大且挖掘行程短，因此转斗挖掘可用于清除障碍或提高生产率。在实际工作中，熟练的液压挖掘机人员可根据实际情况，合理操纵各个液压缸，往往是各液压缸联合工作，实现最有效的挖掘作业。工作装置是工程机械进行生产作业的装置，该装置直接影响到整机的生产率和经济性，因此合理的设计有着重大意义，尤其是土方工程机械，作业过程中动力装置的大部分能量消耗在挖掘土壤上。由于工作装置的重量和成本只占整个机械的很小部分，因此，要降低挖掘土壤的能耗，提高效率，从研究工作装置入手，在通常情况下，仅耗用较少的材料和费用就能明显地提高机械的性能，而机械的结构无需作重大改变。 工程施工对工程机械的工作装置提出的高效、多能，能适应各种作业条件的要求，促使工作装置在结构形式和材料选用上不断提高改进，许多工作装置的操纵已采用液压伺服系统或自动控制，是机械的操纵力大大减小，生产效率利用显著提高，取得了良好的效果。在众多的工作装置中，装载机的工作装置的设计具有一定的典型性二、国内外研究概况及发展趋势（含文献综述）：第一台手动挖掘机问世至今已有130多年的历史，期间经历了由蒸汽驱动半回转挖掘机到电力驱动和内燃机驱动全回转挖掘机、应用机电液一体化技术的全自动液压挖掘机的逐步发展过程。我国的挖掘机生产起步较晚，从1954年抚顺挖掘机厂生产第一台斗容量为1m3的机械式单斗挖掘机至今，大体上经历了测绘仿制、自主研制开发和发展提高等三个阶段。 新中国成立初期，以测绘仿制前苏联20世纪3040年代的W501、W502、W1001、W1002等型机械式单斗挖掘机为主，开始了我国的挖掘机生产历史。1967年开始，我国自主研制液压挖掘机。早期开发成功的产品主要有上海建筑机械厂的WYl00型、贵阳矿山机器厂的W4-60型、合肥矿山机器厂的WY60型挖掘机等。随后又出现了长江挖掘机厂的WYl60型和杭州重型机械厂的WY250型挖掘机等。它们为我国液压挖掘机行业的形成和发展迈出了极其重要的一步。 到20世纪80年代末，我国挖掘机生产厂已有30多家，生产机型达40余种。但总的来说，我国挖掘机生产的批量小、分散，生产工艺及产品质量等与国际先进水平相比，有很大的差距。我国单斗液压挖掘机应向全液压方向发展；斗容量宜控制在0.1-15 m；而对于大型及多斗挖掘机，由于液压元件的制造、装配精度要求高，施工现场维修条件差等，则仍以机械式为主。应着手研究、运用电液控制技术，以实现液压挖掘机操纵的自动化。 改革开放以来，积极引进、消化、吸收国外先进技术，以促进我国挖掘机行业的发展。例如，中国第一拖拉机工程机械公司、广西玉柴股份有限公司、柳州工程机械厂等。这些企业经过几年的努力已达到一定的规模和水平。业内人士指出，我国单斗液压挖掘机应向全液压方向发展；斗容量宜控制在0.1-15 m3；而对于大型及多斗挖掘机，由于液压元件的制造、装配精度要求高，施工现场维修条件差等，则仍以机械式为主。应着手研究、运用电液控制技术，以实现液压挖掘机操纵的自动化。工业发达国家的挖掘机生产较早，法国、德国、美国、俄罗斯、日本等是斗容量3.5-40 m3单斗液压挖掘机的主要生产国，从20世纪80年代开始生产特大型挖掘机。例如，美国马利昂公司生产的斗容量50-150 m3的剥离用挖掘机，斗容量132 m3的步行式拉铲挖掘机；B-E(布比赛路斯一伊利)公司生产的斗容量168.2 m3的步行式拉铲挖掘机，斗容量107 m3的剥离用挖掘机等，是世界上目前最大的挖掘机。 从20世纪后期开始，国际上挖掘机的生产向大型化、微型化、多功能化、专用化和自动化的方向发展。 三、研究内容及实验方案： 主要对由动臂、斗杆、铲斗、连杆机构组成挖掘机工作装置进行设计。具体内容包括以下四部分:1挖掘机工作装置的总体设计。 工作装置是液压挖掘机的主要组成部分之一。因用途不同，工作装置的种类繁多，其中最主要的有反产装置，正铲装置，起重装置和抓斗装置等。而同一种装置也可以有许多种结构形式，有的多达数十种，以适应各种不同的作业条件。主要设计的内容有动臂及斗杆的结构形式。动臂是工作装置中的主要构件，斗杆的结构型式往往取决于动臂的结构型式。反铲动臂可以分为整体式和组合式两类。整体式动臂有直动式和组合式两类。直动式臂结构简单，轻巧，布置紧凑，主要用于悬挂式挖掘机。采用整体式弯动臂有利于得到较大的挖掘深度，它是专用反铲装置的常见形式。整体式弯动臂在弯曲处的结构形式和强度值得注意，近年来悬挂式挖掘机出现了小弯臂的结构形式，是直动臂的改良，动臂的箱型结构可以不用开口，动臂和斗杆油缸及管路的布置也比较方便。整体式动臂结构简单，价廉。刚度相同时结构重量较组合式动臂轻。它的缺点是替换工作装置较少，通用性较差。而组合式动臂有以下优点：（1）工作尺寸和挖掘力可以根据作业条件的变化调整。（2）较合理的满足各种类型作业装置的参数和结构要求。（3）装车运输比较方便。综上选用组合式方案2挖掘机的工作装置详细的机构运动学分析。通过研究平面四杆机构的运动分析与曲柄滑块机构的规律，综合利用作图法与解析法，得出各构件的长度与位置。作图法精度不够，只是一种近似的计算方法。3工作装置各部分的基本尺寸的计算和验证。动臂机构参数的选择 （1）动臂机构的铰点位置的选择。（2）动臂液压缸作用力及闭锁力的确定斗杆机构参数的选择确定斗杆液压缸的铰点位置，行程及力臂比时应该考虑以下因素：（1）保证斗杆液压缸产生足够的斗齿挖掘力。（2）保证斗杆液压缸有必要的闭锁能力。（3）保证斗杆的摆角范围。铲斗机构的参数选择反铲铲斗及机构有四杆机构的，也有六杆机构的。作为机构的杆长参数一般都预先选定。这些参数必须满足以下要求：（1）铲斗的转角范围（2）铲斗机构的载荷分析4工作装置主要部件的结构设计。对由动臂、斗杆、铲斗、连杆机构组成挖掘机工作装置进行设计。二 设计方案反铲装置的合理设计问题至今尚未理想的解决。以往多按经验，采用统计和作图试凑的方法，现在则尽可能采用数解分析方法，并利用电子计算机辅助设计。反铲方案的选择主要依据是设计任务书规定的使用要求，据以决定工作装置是通用或是专用。此处确定为通用工作装置。以反铲为主的通用装置应保证反铲使用要求，并照顾到其他装置的性能。专用装置应根据作业条件决定结构方案，在满足主要作业条件的同时照顾其它条件下的性能。设计方法主要是通过利用经验公式计算设计量，动臂，斗杆，转斗。将计算得到的设计量带入精确的计算公式来验证设计量是否得当。若有出入，则再修改设计量。直到由设计量计算出的最大挖掘深度，挖掘半径。最大卸载高度与给定的参数相吻合。具体步骤如下1 动臂及动臂液压缸的布置确定用组合式或整体式动臂，以及组合式动臂的组合式方式，确定动笔液压缸的布置为悬挂式或是下置式。2 斗杆及斗杆液压缸的布置确定用整体式或组合式斗杆，以及组合式斗杆的组合方式或整体式斗杆是否采用变铰点调节。3 确定动臂与斗杆的长度比，及特性参数K1=L1/L2。对于一定的工作尺寸而言，动臂与斗杆之间的长度比可以在很大的范围内选择。一般当K12时称为长动臂短斗杆方案，当K11.5时属于短动臂长斗杆方案。K1在1.52之间称为中间比例方案。4 确定配套铲斗的种类,斗容量及其主要参数，并考虑铲斗连杆机构传动比是否需要调节。5 根据液压系统的工作压力，流量，系统回路供油方式。工厂制造条件和三化要求等确定各液压缸缸数，缸径，全伸长度与全缩长度之比，考虑到结构尺寸，运动余量，稳定性和构件运动幅度等因素，一般取1=1.61.7，个别情况下因动臂摆角和铰点布置要求可以取1=1.756 斗杆机构参数的选择对于以转斗挖掘为主的中小型反铲，选择斗杆参数时必须注意转斗挖掘时斗杆液压缸的闭锁能力，要求在主要挖掘区内转斗液压缸的挖掘力能得到充分发挥。而斗杆的摆角范围大致在105125之间。在满足工作范围和运输要求的条件下此值尽可能的取得小些，对以斗杆挖掘为主的中型机更应注意到这一点。一般说斗杆愈长，其摆角范围可稍小。当斗杆液压缸和转斗液压缸同时伸出最长时，铲斗前臂与动臂之间的距离应大于10cm。根据斗杆挖掘阻力计算，并参照国内外同类型机器斗杆挖掘力值，按要求的最大挖掘力确定斗杆液压缸的最大作用力臂值l9。斗杆上EFQ取决于结构因素，并考虑到工作范围，一般在130170之间。动臂上DFZ也是结构尺寸，根据结构因素预先估计。斗杆机构参数最后还必须按闭锁性能校核。7.铲斗机构的参数选择反铲铲斗及机构有四杆机构的，也有六杆机构的。转角范围大致为90110，为了满足开挖和最后卸载及运输状态的要求，铲斗总转角往往要达到150180。为了挖掘深沟及垂直侧壁的要求，不使斗底先于斗齿接触土壤，常采用大仰角机构，总转角必须选择适当，不宜过大。设计时还要避免当铲斗液压缸全伸时斗齿尖碰撞斗杆下缘的现象。四、目标、主要特色及工作进度（1） 工作进度表1. 查找资料，外文资料翻译（不少于6000字符），开题报告 第1周-第2周2运动及动力参数计算 第3周-第4周3总装配图设计 第5周-第6周 4. 工作装置各部分基本尺寸设计 第7周-第8周5用UG/proe对系统进行实体建模和设计 第9周-第11周6. 绘制零、部件图 第12周-第13周7. 毕业论文撰写 第14周-第16周8 .答辩准备及论文答辩 第17周-第17周（2） 挖掘机反铲工作装置主要特点：反铲主要用于挖掘停机面以下的土壤。其挖掘轨迹决定于各油缸的运动及其相互配合的情况。当采用动臂油缸工作并进行挖掘时（斗杆油缸和铲斗油缸不工作）可以得到最大的挖掘半径和最长的挖掘行程。此时铲斗的挖掘轨迹是以动臂下铰点为中心，斗齿尖至该铰点的距离为半径而做的圆弧线，其极限挖掘高度和深度（不是最大挖掘深度即圆弧线之起点、终点，分别决定于动臂的最大上倾角和下倾角（动臂对水平线的夹角），也即决定于动臂油缸的行程。由于这种挖掘方式时间长并且由于稳定条件限制挖掘力的发挥，实际工作中基本不采用。在实际挖掘工作过程中，往往需要采用各种油缸的联合工作。如当挖掘基坑时由于挖掘深度较大，并要求有较陡而平整的基坑时，则采用动臂与斗杆两种油缸同时工作当挖掘坑底，挖掘行程将结束为加速将铲斗装满土，以及挖掘过程需要改变铲斗与切削角等情况下，则要求采用斗杆与铲斗油缸共同工作。显然此时挖掘机的挖掘轨迹是由相应油缸分别工作时的轨迹组合而成。当然，这种动作能否实现决定于液压系统的设计络图，即挖掘机在任意正常工作周位置时，控制到的工作范围，图上各控制尺寸即液压挖掘机的工作尺寸。对于反铲装置的只要工作尺寸为最大挖掘深度和最大挖掘半径，包络图中可能有部分区间靠近甚至深入到挖掘机停机点以下，这一范围的土壤虽然能挖掘但可能引起土壤的崩塌而影响机械的稳定性和安全性，除有条件的挖沟作业一般不使用。故有的挖掘机工作尺寸图上标明有效的工作范围，或以虚线表明此段的挖掘轨迹如图所示。 图一挖掘机反铲装置的最大挖掘力决定于液压系统的工作压力，油缸尺寸，以及油缸间作用力影响（斗杆，动臂油缸的闭锁压力及力臂）外。还决定与整机的稳定和地面的附着情况。反铲挖掘机速度结构尺寸已定的条件下决定于液压系统对工作油缸的供油量，对动臂油缸和斗杆油缸为提高其单独工作的挖掘速度，在液压系统中可采用合流供油措施来保证。液压反铲都采用转斗卸土，卸载较准确，平稳，便于装车工作。 五、参考文献1成大先主编.机械设计手册(第三版,第三卷),第十四篇M.化学工业出版社,1992年2付 越，邓子龙. 基于ProE的液压挖掘机反铲工作装置运动仿真，辽宁石油化工大学学报.2007.6.pp50-53.3 邓子龙，刘杰，高财禄等.挖掘机铲斗结构优化.机械与电子.2009.1.pp13-16.4 李滨城，何允纪.液压挖掘机反铲工作装置运动的模拟.华东船舶工业学院学报.1995.6.pp74-805 刘玉强，王学军.液压挖掘机反铲工作装置优化设计，机械产品与科技.1997.1.pp12-156 Yu. I. Berezhnoi and Yu. A. Potapov. Method for the intermittent optimization of the working of a block by a rotary excavator Refractories and Industrial Ceramics, 1982, Volume 23, Numbers 3-4, Pages 183-186 MINING MECHANIZATION AND AUTOMATION GEOMETRY OF THE WORKING PART OF AN EXCAVATOR TOOTH V. A. Polovinko and A. I. Fedulov UDC 621.879.3 Studies of excavator tooth wear kinetics conducted earlier by the present authors 1, 2 showed that the main factor controlling wear platform dynamics is the physical-mechanical property of the rock. Wear platforms evolve in two stages. Tooth wear acquired during the critical stage 2 has no significant influence on excavator performance in the mining and geologic conditions typical for the northeastern regions of Russia. Cutting elements can continue to be used up to the maximum permissible wear level specified by the manufacturer. In this respect, intensive wear during initial stages apparently reflects some design imperfection rather than the effects of the work adjustment process. Investigators have studied the causes and consequences of intense wear of excavator teeth, but there are still no basic criteria upon which to formulate general principles so as to improve the wear resistance of cutting elements as determined by their design 3-5. An efficient way to raise the wear resistance of an excavator tooth is to devise the design parameters of the working component so as to ensure classical single-stage wear, bypassing the critical (pseudoadjustment) phase. We developed a new excavator tooth design which features heightened wear resistance. The outline of the working component of the tooth and its dimensions were developed with due regard for the main characteristic points of the wear resistance curves of mass-produced wedge-shaped teeth. To attain a linear behavior for the wear process of such teeth with a rate equal to or less than what is observed during the second stage of wear with mass- produced teeth, we specified the design parameters corresponding to the beginning of the second phase, where the specific pressure from the standard force of the thrust mechanism drops to 10-12 MPa. Figure 1 plots pressure variations on the wear platforms of teeth of buckets used in common quarry excavators according to the following expression: P1 where Up is the width of the wear platform; P1 is the rated force of the thrust mechanism; D and i are the length of the tooth cutting edge and the number of teeth on the bucket, respectively. The curves show that there are certain pressure regions on wear platforms where rock resistance to teeth is equal to or greater than the force developed by the thrust mechanism. This loading pattern for cutting elements is observed on monolithic strong (e.g., permafrost) rocks. On the other hand, some materials resist cutting with a much weaker strength than the force developed by this thrust mechanism. To estimate the specific pressures formed when cutting elements interact with these materials, we plotted curves 1-4 by computing the pressure on the wear platforms of an tKG-5A excavator tooth at 0.8, 0.4, 0.2, and 0.1 of the rated thrust force. On weak rocks the pressure variation pattern on the wear platform is the same, but the pressures and dimensions for the worn portion of teeth after the beginning of the second stage may be much smaller (sometimes by a considerable factor). This is clearly seen in Fig. 1. Zone I, crossing the curves, defines the parameters of the onset of the second stage of wear for teeth of different excavators and for different rock strengths (curves 1-4). For IKG-5A excavator teeth the starting point of the second wear stage obtained experimentally lies in zone I and corresponds to a pressure of P = 10-12 MPa and a wear platform width of Utc r = 45 mm. Institute of Mining, Siberian Branch, Russian Academy of Sciences, Novosibirsk. Translated from Fiziko- Tekhnicheskie Problemy Razrabotki Poleznykh Iskopaemykh, No. 2, pp. 16-23, March-April, 1993. Original article submit- ted November 4, 1992. 1062-7391/93/2902-0115512.50 1993 Plenum Publishing Corporation 115 0 MPa tl/( . O EKG-20 EKG-12,5 EKG-$I 5A Up Up, mm Fig. 1 Fig. 2 Fig. 1. Pressure variation as a function of wear platform size (1-4 - theoretical pressure curves on an 1KG- 5A excavator tooth wear platform when working rocks with resistance 0.8, 0.4, 0.2, and 0.1 of standard thrust force). Fig. 2. Working part of a cutting element with wedge angle 180 (1 cutting edge with area So; b - edge width; D - length; 2 - wear platform surface area Sp2; 1, - wear platform slope angle. At a given size of the working part of the tool, the stage of critical wear or pseudoadjustment is virtually absent on rocks and grounds with low strength, while tools experience intense two-stage wear on strong/hard rocks. In different mining and geologic conditions, it is obviously convenient to work with interchangeable tools. It is currently impossible to control the force parameters on the working element of an excavator. The operator observes the work of the machine visually, watching its motion and bucket filling. The loads acting upon working elements and teeth thus depend not only on rock resistance to cutting, but largely on operator skill and experience. Art efficient and rational approach to devising working tooth component parameters is to consider the power of excavator drives. The area of the cutting edge for a rectangular cutting profile with a 180 sharpening can be calculated from the pressure on the wear platform (see Fig. 2) corresponding to the onset of the second wear stage: P = PI Sp2. i where P is the pressure on the wear platform when platform dimensions correspond to the beginning of the second stage; P1 is the rated thrust force of the excavator (vertical component of the cutting force); Sp2 is the wear platform area at the second stage onset; i is the number of teeth on the excavator bucket. The wear platform is defined in terms of the cutting edge area as S O SP2 = sinT where 3 is the wear platform slope angle relative to the back facet of the cutting profile; S o is cutting edge area. The pressure on the wear platform can be expressed as p = .Px sin ? So- The area of the cutting edge which provides the desired wear pattern for the cutting element is defined from the same formula: So - P1 sin 7 P-i 116 / 2? -7 r j U, mm Ve z1V j I V, 1,000 m 3 Fig. 3 Fig. 4 Fig. 3. Cutting elements with heightened wear resistance (Ucr - linear wear corresponding to first critical stage; b b and D b - basic width and length of cutting edge of wedge tooth; D - calculated length of cutting edge). Fig. 4. Design-controlled wear resistance of wedge-shaped cutting elements (1, 2 varia- tion of linear wear for a tooth with an expanded part and a standard tooth, respectively; Umax - maximum permissible wear; &V - increased operation resource of new tooth design. Considering that the cutting edge area is linked to the wear platform by the preceding relation, we can formulate simple technological conditions for improving the design of the working component of standard wedge-shaped teeth in terms of optimal length of the tooth cutting edge as D = Sp2.s_ni ? b where D is the optimal cutting edge length which provides steady single-stage wear of cutting elements; b is the actual (basic) width of the cutting edge of mass-produced wedge teeth; Sp2 is the area of the platform corresponding to the onset of steady wear; and 3 is the angle of the slope of the wear platform with respect to tooth longitudinal axis. Figure 3 offers technological concepts for reduction of cutting element wear dynamics based on mass-produced wedge-shaped teeth. The length of the expanded part of a tooth (D) should be not less than critical linear wear Ucr. After the expanded part is worn off, a tooth acquires the natural size of the platform corresponding to the second stage of steady wear. This design wears according to a linear relationship (Fig. 4) with an intensity equal to *._hat of the second stage of wear of mass-produced teeth (parallel portions of plots). After attaining maximum wear, teeth would have extended service life, expressed in an increased volume of excavated rock (AV). We should pay special attention to creating teeth with heightened wear resistance without modifying the basic dimensions or shape of the working component. This is important, because this form is easier and less expensive to manufac- ture. We developed the universal geometry for the working part of an excavator tooth based on calculations of the optimal width of the cutting edge while retaining the main dimensions of standard teeth designs.* The tooth with the new working component geometry (Fig. 5) has cutting edge 1, linear segment of back face 2, and curvilinear part 3. The front face is formed of two linear segments 4 and 5. The linear segment of back face 2 is parallel to tooth longitudinal axis 6, situated at distance I from the axis 6. The plane of the cutting edge is situated at an angle greater than 90 to the cutting plane. This helps form a steady compaction core on the plane Of the cutting edge, which partly protects it from wear. The cutting edge width is found from an empiric relationship: *We took the tooth design developed by the Institute of Heavy Machinery (Uralmash Production Association) for the basic prototype. !17 7 7 6 Fig. 5. Design of the working part of a tooth with optimal parameters (1 - cutting edge; 2 - linear portion of the back facet; 3 - curvilinear back facet; 4, 5 - segments of the front facet; 6 - longitudinal tooth axis; 7- wear platform; b = cuing edge width; a 1 = initial cutting angle; / = wedge angle; f = distance between wedge angle vertex and cutting edge; A and B = dimensions of linear segments of front and back facets, respectively; I = displacement of back facet segment from tooth axis; r = wear platform slope angle. b PI sin P,D. where b is an efficient width of the cutting edge; P1 is the excavator thrust force, which consists of the weight of bucket and the stick, and the force developed by the thrust mechanism; 3 is the slope angle of the wear platform relative to the tooth axis (or the linear segment of the back facet); D is the length of the cutting edge; P is the pressure on the wear platform at the beginning of the second stage; and i is the number of the teeth on the bucket. Cutting edge 1 should be at distance PI sin from the vertex of the wedge angle, where is the wedge angle of the working part of the tooth. Literary data indicate that a change of the cutting angle (more precisely, the back angle, which depends on the cutting angle) greatly affects the intrusion force of cutting elements. When the back angle of a tooth is increased, the energy capacity of its intrusion into the ground tends to decrease 7. We formulated the new tooth geometry taking this factor into account. Accordingly, linear segment 2 or back facet 3 is parallel to tooth axis 6, which allowed us to increase the back angle by a factor of 2.0-2.5 compared with the mass-produced model. To reduce the wear of the horizontal component of the cutting parameter of the excavator bucket, we shifted segment 2 of the back facet (and thus cutting edge 1) by value F from the tooth axis. This position of the elements of the tooth working component relative to the bucket cutting edge reduces the rate of wear because the distance between the tooth cutting plane and the bucket edge cutting plane is increased by 70-80% for a given length of the working tooth part protruding beyond the bucket. For this tooth design, we defined the relationship which can be used to calculate the dimensions of the working elements (Table 1). The length of the cutting edge (D) and the number of teeth (i)are chosen depending on the design of the excavator working element and the general machine specifications. Figure 6 shows theoretical curves of the formation of wear platform dimensions as a function of linear wear for an tKG-5A excavator. We can see that at zero wear the design with efficient parameters has a wear platform of = 50 ram. The design produced by the Uralmash Production Association attains the desired dimensions only after significant linear wear Ucr. With further wear (Fig. 6, zone II) the wear platform evolves less rapidly and the wear rate is approximately equal to that of 118 TABLE 1 Dimensions of working Calcualted parameters of IKG-5A Parameter component dements excavator tooth (4, 6) Cutting edge width : b = 28 mm Distance between vertex of tooth wedge angle and cutting edge Wedge angle of working part of tooth Wear platform slope angle Length of linear segment of back facet Length of linear segment of front facet Displacement of linear segment of back facet from tooth axis P-sin ? b- P.D.i P, .sin ? 1- P-D.i.tg = 33 . 37 7=40 . 45 . B= (4,5-5,8)b A= (3,0-4,0)b F= (2,0-3,5)b /-40 mm 13 = 35 ? = 40 B = t50 mm A = 100 :ram I = 70 mm the tooth made by Uralmash. In other words, during wear stages, wear resistance in zones I and II accumulates a certain reserve because the wear platform of a new tooth develops more rapidly in the initial stages (up to 50 mm). Subsequently, a change in wear platform size has no significant influence on wear rate. The principles for working component development and parameter evaluation which make it possible to predict and control the wear dynamics of cutting elements were tested in real industrial conditions. The Orotukan Mining Machine Factory of the Severovostokzoloto Gold Mining Production Association manufactured a test batch of excavator teeth follow- ing the new design. They were tested at the Yagodnin, Berelekh, and other mining enterprises. Valuable test results were obtained at the Korba facility at the Burkandya Mine (Berelekh Enterprise), where teeth were used to excavate highly abrasive frozen rocks consisting of granite, sandstone, and clay schist fragments at surrounding air temperatures of -45 C. Analysis of the design-related wear resistance of these experimental teeth indicated that, with the new working component geometry they experience single-stage wear with a resistance at least 40% higher than that of standard wedge teeth (Fig. 7). Experimental data were analyzed to define linear wear U as a function of work output V (thousand m 3) for new teeth. It was expressed by a first-degree regression equation: U = 8.57 + 5.g2V with a correlation coefficient of 0.980. The test confirmed the basic design principles of cutting elements based on an empiric relationship of the working element geometry with pressure on the wear platform expressed in terms of the force characteris- tics of excavator working components. Cutting edge width can be defined from b = Up= -sin,?, (1) where . is the wear is the wear platform for the onset of steady-state tooth wear, which is defined from a chart (see Fig. 1); 3 is the wear platform slope angle. Using the empiric relationship (see Table 1) and rearranging it with substitution of numeric values of P based on tests, we can estimate the cutting edge width as zD 1 b = 0,0536 D.i (2) which guarantees single-stage wear in any operation conditions. This is achieved because the expression contains a constant coefficient (pressure P = 12 MPa) which serves as the main criterion. At the beginning of linear wear, a specific pressure of 12 MPa operates on the wear platform of a cutting element designed according to this formula. This analytic technique is simple and reliable because it provides a wear resistance margin. This is important for excavator teeth used to cut bedrock, which are usually rapidly blunted. The sharpness of teeth in this case is of little practical importance because the wear intensity is extremely high. A similar positive effect of this method can be expected for teeth used to cut frozen rocks, which are characterized by strong resistance to intrusion leading to back facet wear and formation of a wear platform that sharpens the teeth (reduces cutting edge). 119 U, U 160 .m :00- a b 80- !- ee o/ o o 4 do 120 , mill 1 7 80- 40- -& lk g4 jk 7 X l,O00 m 3 Fig. 6 Fig. 7 Fig. 6. Variation of wear platform dimensions as a function of linear wear (a - formation of wear platform of cutting element with improved parame- ters of working part; b - formation of wear platform of Uralmash tooth design; I, II - development zones of wear platform before onset of linear wear (I) and during linear wear (II). Fig. 7. Design-controlled wear resistance of KG-5A excavator tooth (wear dynamics of new tooth design (a) and standard wedge tooth (b); 1, 2 - l