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液压挖掘机反铲工作装置设计【优秀含6张CAD图纸+proe三维建模+挖掘机工程机械全套课程毕业设计】

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asm.asm.10

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b.prt.4

b1.bak

body.prt.14

body1.bak

c.prt.1

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std.out

syscol.scl

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弯动臂-A3.dwg

斗杆-A3.dwg

斗齿-A3.dwg

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Geometry of the working part of__ an excavator tooth.pdf

学士学位论文原创性声明.doc

摘要.doc

毕业论文开题报告.doc

毕业设计(1).doc

液压挖掘机反铲装置设计说明书.doc

翻译.docx

翻译.wps

翻译封面.doc

proe三维

弯动臂-A3.dwg

斗杆-A3.dwg

斗齿-A3.dwg

液压挖掘机工作装置-A1.dwg

液压挖掘机总图-A0.dwg

设计文档

铲斗-A2.dwg

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

Ⅳ 、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.pp74-80

[5] 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.pp12-15

[6] 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

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。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2～3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；反铲机构；设计

ABSTRACT

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. 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

1　绪论1

1.1液压挖掘机的工作特点和基本类型1

1.1.1液压挖掘机的主要优点1

1.1.2液压挖掘机的基本类型及主要特点2

1.2反铲装置的工作原理2

2　总体设计方案4

2.1工作装置设计方案原则4

2.2液压系统设计方案原则（总体） 4

2.2.1对液压系统作业动作要求 4

2.2.2对液压系统基本的要求 5

3挖掘机工作装置设计6

3.1确定动臂的结构形式6

3.2动臂、铲斗机构参数的选择6

3.2.1反铲装置总体方案的选择6

3.2.2铲斗参数的选择 7

3.2.3 动臂机构参数的选择 8

4 液压挖掘机工作装置运动仿真 13

4.1模型建立13

4.2构件运动配装15

4.3构件运动仿真17

结论21

致谢22

参考文献23

1　绪论

　液压挖掘机是在机械传动挖掘机的基础上发展起来的。它的工作过程是以铲斗的切削刃切削土壤，铲斗装满后提升、回转至卸土位置，卸空后的铲斗再回到挖掘位置并开始下一次的作业。因此，液压挖掘机是一种周期作业的土方机械。

　　液压挖掘机与机械传动挖掘机一样，在工业与民用建筑、交通运输、水利施工、露天采矿及现代化军事工程中都有着广泛的应用，是各种土石方施工中不可缺少的一种重要机械设备。

　　在建筑工程中，可用来挖掘苦坑、排水沟，拆除旧有建筑物，平整场地等。更换工作装置后，可进行装卸、安装、打桩和拔除树根等作业。

　　在水利施工中，可用来开挖水库、运河、水电站堤坝的基坑、排水或灌溉的沟渠，疏浚和挖深原有河道等。

　　在铁路、公路建设中，用来挖掘土方、建筑路基、平整地面和开挖路旁排水沟等。

　　在石油、电力、通信业的基础建设及市政建设中，用来挖掘电缆沟和管道等。

　　在露天采矿场上，可用来剥离矿石或煤，也可用来进行堆弃、装载和钻孔等作业。

所以，液压挖掘机作为工程机械的一个重要品种，对于减轻工人繁重的体力劳动，提高施工机械化水平，加快施工进度，促进各项建设事业的发展，都　起着很大的作用。据建筑施工部门统计，一台斗容量1.0m3的液压挖掘机挖掘Ⅰ～Ⅳ级土壤埋，每班生产率大约相当于300～400个工人一天的工作量。因此，大力发展液压挖掘机，对于提高劳动生产率和加速国民经济的发展具有重要意义。

参考文献

［1］曹善华、余涵.单斗液压挖掘机.北京：中国建筑工业出版社，1980

［2］孔德文、赵克利.液压挖掘机.北京：化学工业出版社，2007

［3］高衡、张全根.液压挖掘机.北京：中国建筑工业出版社，1981

［4］阎书文.机械式液压挖掘机.北京：机械工业出版社，1982

［5］天津工程机械研究所.单斗液压挖掘机.北京：中国建筑工业出版社，1976

［6］何存兴.液压传动与气压传动.华中科技大学出版社，2000

［7］张铁.液压挖掘机结构、原理及使用.东营：石油大学出版社，2002

［8］黄宗益，王康.液压挖掘节能控制.建筑机械，1997

［9］张平格.液压传动与控制.北京：冶金工业出版社，2004

［10］成大先.机械设计手册.北京：化学工业出版社，2002

［11］吴相宪、王正为、黄玉堂.实用机械设计手册。徐州：中国矿业大学出版社，1993

［12］唐大放、冯晓宁、杨现卿.机械设计工程学.中国矿业大学出版社，2001

［13］李壮云.中国机械设计大典.江西科学技术出版社，2001

［14］唐经世、高车安.工程机械.北京：中国铁道出版社，1996

［15］周士昌.液压系统设计图集.北京：机械工业出版社，2003

［16］杜迪生、张永惠.挖掘机电气传动与故障诊断.北京：冶金工业出版社，1994

［17］张玉川.进口液压挖掘机国产化改造.成都：西南交通大学出版社，1999

［18］R.N.Hancox,Hydraulic System for Excamator,U.S.Patent 3406850.

OCT.22,1968

［19］R.K.Tessmann,I.T.Hong,Hydraulic Pump Performance as a Function

Of Speed and Pressure,SAE961741

[20] 　张华,郭荣春,周进. 挖掘机动臂在Pro/ E 中的动态模拟与分析[J ] . 农业装备与车辆工程,2005.

[21] 　王志利,韩刚. Pro/ EN GINEER 在挖掘机设计中的应用[J ] . 起重运输机械,2005

[23] 　孙印杰,田效伍,郑延斌. 野火中文版Pro/ Engineer 基础与实例教程[M] . 北京:电子工业出版社,2004.

[24] 　方建军,刘仕良. 机械动态仿真与工程分析———Pro/ Engineer Wildfire 工程应用[M] . 北京:化学工业出版社,2004 .

[25] 　范进桢,张宝忠,秦贵林. 挖掘机的运动学分析[J ] . 煤矿机械,2004

[26] 　朱向哲,林伟,谢禹钧. 双螺杆非啮合螺纹元件错列角对挤出特性的影响[J ] . 石油化工高等学校学报,2005 ,18

[27] 　祝凌云,李斌. Pro/ EN GINEER 运动仿真和有限元分析[M] . 北京:人民邮电出版社,2004.

[28] 　郭卫,杨武成,张传伟. 基于Pro/ E 的液压挖掘机工作装置运动仿真[J ] . 建筑机械

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内容简介：: 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 nts0 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 nts/ 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 nts7 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 ntsTABLE 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 ntsU, 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 - linear wear and wear platform formation, respectively). TABLE 2 Characteristic Excavator parameters: thurst mechansm force, MN weight of bucket with stick, tons total pressure force, MN number of teeth per bucket, units . Conventional teeth: basic length 0f. e2tt.iflg.-edge_, mm basic width of cutting edge, ram wedge-shaped teeth with yedge angle 20-30 teeth with curved back and front facets New design: calculated width of cutting edge, mm according to (i) according to (2) KG:5A(4,6) 0,205 20,0 0,405 Excavator EKG4I gKG-12,5 0,37 0,60 26,5 45,0 0,635 1,05 IKG-20 0,70 8i,5 i,565 t50 i2 (t6) 28 28 180 6(0) 34 37 t80 20 44 63 220 45, * 44 64, *Wedge angle of working component 38 . The teeth parameters formulated with our method and experimentally tested can be compared with currently used models on quarry excavators (Table 2). 120 ntsThe data in Table 2 dearly show that most current designs do not meet the needs of frozen rock excavation. Our studies and calculations confirm this observation. Standard teeth with increased cutting edge dimensions (given in parenthe-

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