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反铲式单斗液压挖掘机工作装置设计及其运动分析设计【9张CAD图纸+PDF图】

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目 录前言 1一、绪论2(一)国内外研究状况2(二)论文构成及研究内容2二、总体方案设计3(一)工作装置构成3(二)动臂及斗杆的结构形式5(三)动臂油缸与铲斗油缸的布置5(四)铲斗与铲斗油缸的连接方式5(五)铲斗的结构选择6(六)原始几何参数的确定7三、工作装置运动学分析8(一)动臂运动分析8(二)斗杆的运动分析10(三)铲斗的运动分析11(四)特殊工作位置计算15四、挖掘阻力分析18(一)转斗挖掘阻力计算18(二)斗杆挖掘阻力计算18五、基本尺寸的确定 20(一)斗形参数的确定20(二)动臂机构参数的选择201、 1与A点坐标的选取202、 l1与l2的选择203、 l41与l42的计算214、 l5的计算 21(三)动臂机构基本参数的校核 231、 动臂机构闭锁力的校核232、 满斗处于最大挖掘半径时动臂油缸提升力矩的校核253、 满斗处于最大高度时,动臂提升力矩的校核 26(四)斗杆机构基本参数的选择27(五)铲斗机构基本参数的选择281、 转角范围282、 铲斗机构其它基本参数的计算28六、工作装置结构设计 30(一)斗杆的结构设计301、斗杆的受力分析 302、斗杆内力图的绘制 353、 结构尺寸的计算37(二)动臂结构设计391、危险工况受力分析 422、内力图和弯矩图的求解 433、 结构尺寸的计算45(三)铲斗的设计471、铲斗斗形尺寸的设计 472、铲斗斗齿的结构计算 473、 铲斗的绘制48七、销轴与衬套的设计 49(一)销轴的设计49(二)销轴用螺栓的设计49(三)衬套的设计49八、总结50九、参考文献 51十、致谢52附件一 外文翻译 53反铲式单斗液压挖掘机工作装置设计及其运动分析设计引 言挖掘机在国民经济建设的许多行业被广泛地采用,如工业与民用建筑、交通运输、水利电气工程、农田改造、矿山采掘以及现代化军事工程等等行业的机械化施工中。据统计,一般工程施工中约有60%的土方量、露天矿山中80%的剥离量和采掘量是用挖掘机完成的。随着我国基础设施建设的深入和在建设中挖掘机的广泛应用,挖掘机市场有着广阔的发展空间,因此发展满足我国国情所需要的挖掘机是十分必要的。而工作装置作为挖掘机的重要组成部分,对其研究和控制是对整机开发的基础。反铲式单斗液压挖掘机工作装置是一个较复杂的空间机构,国内外对其运动分析、机构和结构参数优化设计方面都作了较深入的研究,具体的设计特别是中型挖掘机的设计已经趋于成熟。关于反铲式单斗液压挖掘机的相关文献也很多,这些文献从不同侧面对工作装置的设计进行了论述。而笔者的设计知识和水平还只是一个学步的孩子,进行本课题的设计是为对挖掘机的工作装置设计有一些大体的认识,掌握实际工程设计的流程、方法,巩固所学的知识和提高设计能力。一、绪论(一)国内外研究状况当前,国际上挖掘机的生产正向大型化、微型化、多能化和专用化的方向发展。国外挖掘机行业重视采用新技术、新工艺、新结构和新材料,加快了向标准化、系列化、通用化发展的步伐。我国己经形成了挖掘机的系列化生产,近年来还开发了许多新产品,引进了国外的一些先进的生产率较高的挖掘机型号。由于使用性能、技术指标和经济指标上的优越,世界上许多国家,特别是工业发达国家,都在大力发展单斗液压挖掘机。目前,单斗液压挖掘机的发展着眼于动力和传动系统的改进以达到高效节能;应用范围不断扩大,成本不断降低,向标准化、模块化发展,以提高零部件、配件的可靠性,从而保证整机的可靠性;电子计算机监测与控制,实现机电一体化;提高机械作业性能,降低噪音,减少停机维修时间,提高适应能力,消除公害,纵观未来,单斗液压挖掘机有以下的趋势:1、向大型化发展的同时向微型化发展。2、更为普遍地采用节能技术。3、不断提高可靠性和使用寿命。4、工作装置结构不断改进,工作范围不断扩大。5、由内燃机驱动向电力驱动发展。6、液压系统不断改进,液压元件不断更新。7、应用微电子、气、液等机电一体化综合技术。8、增大铲斗容量,加大功率,提高生产效率。9、人机工程学在设计中的充分利用。(二)论文构成及研究内容本论文主要对由动臂、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。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。5、 销轴的设计及螺栓等标准件进行选型。二、总体方案设计(一)工作装置构成1-斗杆油缸;2- 动臂; 3-油管; 4-动臂油缸; 5-铲斗; 6-斗齿; 7-侧板;8-连杆; 9-曲柄: 10-铲斗油缸; 11-斗杆图2.1 工作装置组成图 图2.1为液压挖掘机工作装置基本组成及传动示意图,如图所示反铲工作装置由铲斗5、连杆9、斗杆11、动臂2、相应的三组液压缸1, 4,10等组成。动臂下铰点铰接在转台上,通过动臂缸的伸缩,使动臂连同整个工作装置绕动臂下铰点转动。依靠斗杆缸使斗杆绕动臂的上铰点转动,而铲斗铰接于斗杆前端,通过铲斗缸和连杆则使铲斗绕斗杆前铰点转动。挖掘作业时,接通回转马达、转动转台,使工作装置转到挖掘位置,同时操纵动臂缸小腔进油使液压缸回缩,动臂下降至铲斗触地后再操纵斗杆缸或铲斗缸,液压缸大腔进油而伸长,使铲斗进行挖掘和装载工作。铲斗装满后,铲斗缸和斗杆缸停动并操纵动臂缸大腔进油,使动臂抬起,随即接通回转马达,使工作装置转到卸载位置,再操纵铲斗缸或斗杆缸回缩,使铲斗翻转进行卸土。卸完后,工作装置再转至挖掘位置进行第二次挖掘循环。在实际挖掘作业中,由于土质情况、挖掘面条件以及挖掘机液压系统的不同,反铲装置三种液压缸在挖掘循环中的动作配合可以是多样的、随机的。上述过程仅为一般的理想过程。挖掘机工作装置的大臂与斗杆是变截面的箱梁结构,铲斗是由厚度薄的钢板焊接而成。各油缸可看作是只承受拉压载荷的杆。根据以上特征,可以对工作装置进行适当简化处理。则可知单斗液压挖掘机的工作装置可以看成是由动臂、斗杆、铲斗、动臂油缸、斗杆油缸、铲斗油缸及连杆机构组成的具有三自由度的六杆机构,处理的具体简图如2.2所示。进一步简化得图如2.3所示。图2.2 工作装置结构简图1-铲斗;2-连杆;3-斗杆;4-动臂;5-铲斗油缸;6-斗杆油缸图2.3 工作装置结构简化图挖掘机的工作装置经上面的简化后实质是一组平面连杆机构,自由度是3,即工作装置的几何位置由动臂油缸长度L1、斗杆油缸长度L2、铲斗油缸长度L3决定,当L1、L2、L3为某一确定的值时,工作装置的位置也就能够确定。(二)动臂及斗杆的结构形式动臂采用整体式弯动臂,这种结构形式在小型挖掘机中应用较为广泛。其结构简单、价廉,刚度相同时结构重量较组合式动臂轻,且有利于得到较大的挖掘深度。斗杆也有整体式和组合式两种,大多数挖掘机采用整体式斗杆。在本设计中由于不需要调节斗杆的长度,故也采用整体式斗杆。(三)动臂油缸与铲斗油缸的布置动臂油缸装在动臂的前下方,动臂的下支承点(即动臂与转台的铰点)设在转台回转中心之前并稍高于转台平面,这样的布置有利于反铲的挖掘深度。大部分中小型液压挖掘机以反铲作业为主,常采用动臂支点靠前布置的方案。油缸活塞杆端部与动臂的铰点设在动臂箱体下底板的凸缘上,虽然这样会影响动臂的下降幅度,但不会削弱动臂的结构强度,而且使动臂的受力更加合理。对于斗容量为0.25 m3的小型液压挖掘机,单只动臂液压缸即可满足工作要求。具体结构如图2.2所示。(四)铲斗与铲斗油缸的连接方式本方案中采用六连杆的布置方式,相比四连杆布置方式而言在相同的铲斗油缸行程下能得到较大的铲斗转角,改善了机构的传动特性。该布置中1杆与2杆的铰接位置虽然使铲斗的转角减少但保证能得到足够大的铲斗平均挖掘力。如图2.4所示。2331-斗杆; 2-连杆机构; 3-铲斗图2.4 铲斗连接布置示意图(五)铲斗的结构选择铲斗结构形状和参数的合理选择对挖掘机的作业效果影响很大,合适的铲斗应满足以下要求:1、有利于物料的自由流动。铲斗内壁不宜设置横向凸缘、棱角等。斗底的纵向剖面形状要适合于各种物料的运动规律。2、要使物料易于卸尽。3、为使装进铲斗的物料不易于卸出,铲斗的宽度与物料的粒径之比应大于4,大于50时,颗粒尺寸不考虑,视物料为均质。综上考虑,选用小型挖掘机常用的铲斗结构,基本结构如图2.5所示。图2.5 铲斗斗齿的安装连接采用橡胶卡销式,结构示意图如2.6所示。1-卡销 ;2 橡胶卡销;3 齿座; 4斗齿图2.6 卡销式斗齿结构示意图(六) 原始几何参数的确定1、动臂与斗杆的长度比K1由于所设计的挖掘机适用性较强,作业对象明确,一般不替换工作装置,故取中间比例方案,K1取在1.52.0之间。考虑到K1值大,工作装置结构重心离机体近。初步选取K1=2,即l1 / l2=2。2、铲斗斗容与主参数的选择斗容量在任务书中已经给出:q =0.25 m3按经验公式和比拟法初选:l3=900mm,铲斗平均宽度B=800mm,铲斗切削半径R= l3=900mm,铲斗装满转角。3、工作装置液压系统主参数的初步选择各工作油缸的缸径选择要考虑到液压系统的工作压力和“三化“要求。初选动臂油缸内径D1=125mm,活塞杆的直径d1=80mm。斗杆油缸的内径D2=90mm,活塞杆的直径d2=63mm。铲斗油缸的内径D3=100mm,活塞杆的直径d3=70mm。按经验公式初选各油缸全伸长度与全缩长度之比:1=2=3=1.6。参照任务书的要求选择工作装置液压系统的工作压力P=20MPa,闭锁压力Pg=21MPa。三、工作装置运动学分析(一) 动臂运动分析动臂油缸的最短长度;动臂油缸的伸出的最大长度;A:动臂油缸的下铰点;B:动臂油缸的上铰点;C:动臂的下铰点.图3.1 动臂摆角范围计算简图动臂摆角1是L1的函数。动臂上任意一点在任一时刻的坐标值也都是L1的函数。如图3.1所示,图中动臂油缸的最短长度;动臂油缸的伸出的最大长度;动臂油缸两铰点分别与动臂下铰点连线夹角的最小值;动臂油缸两铰点分别与动臂下铰点连线夹角的最大值;A:动臂油缸的下铰点;B:动臂油缸的上铰点;C:动臂的下铰点。则有:在三角形ABC中: (3-1)图3.2 F、C点坐标计算简图在三角形BCF中: (3-2)由图3.2所示的几何关系,可得到21的表达式: (3-3)当F点在水平线CU之下时21为负,否则为正。F点的坐标为 XF = l30+l1cos21 YF = l30+l1sin21 (3-4)C点的坐标为 YC = YA+l5sin11 (3-5)动臂油缸的力臂e1 (3-6)显然动臂油缸的最大作用力臂e1max= l5(二)斗杆的运动分析如下图3.3所示,D点为斗杆油缸与动臂的铰点点,F点为动臂与斗杆的铰点,E点为斗杆油缸与斗杆的铰点。斗杆的位置参数是l2,这里只讨论斗杆相对于动臂的运动,即只考虑L2的影响。D-斗杆油缸与动臂的铰点点; F-动臂与斗杆的铰点;E-斗杆油缸与斗杆的铰点; 2-斗杆摆角.图3.3 斗杆机构摆角计算简图在三角形DEF中 (3-7)由上图的几何关系知斗杆相对于动臂的摆角范围2max2max =2 max-2min (3-8)则斗杆的作用力臂 (3-9)显然斗杆的最大作用力臂e2max = l9,此时。(三)铲斗的运动分析铲斗相对于XOY坐标系的运动是L1、L2、L3的函数,现讨论铲斗相对于斗杆的运动,如图3-4所示,G点为铲斗油缸与斗杆的铰点,F点为斗杆与动臂的铰点Q点为铲斗与斗杆的铰点,v点为铲斗的斗齿尖点,K点为连杆与铲斗的饺点,N点为曲柄与斗杆的铰点,M点为铲斗油缸与曲柄的铰点,H点为曲柄与连杆的铰点。图3.4 铲斗连杆机构传动比计算简图1、铲斗连杆机构传动比i利用图3.4,可以求得以下参数:在三角形HGN中32 = GMN = - MNG - MGN = -22-30 (3-10)在三角形HNQ中 (3-11)在三角形QHK中 (3-12)在四边形KHNQ中NHK=NHQ+QHK (3-13)铲斗油缸对N点的作用力臂r1 (3-14)连杆HK对N点的作用力臂r2r2 = l13Sin NHK 连杆HK对Q点的作用力臂r3 (3-15)连杆机构的总传动比i (3-16)显然3-17式中可知,i是铲斗油缸长度L3的函数,用L3min代入可得初传动比i0,L3max代入可得终传动比iz。2、铲斗相对于斗杆的摆角3铲斗的瞬时位置转角为 (3-17)其中,在三角形NFQ中 (3-18)当铲斗油缸长度L3分别取L3max和L3min时,可分别求得铲斗的最大和最小转角3max和3min,于是得铲斗的摆角范围: 3 = 3max-3min (3-19)3、斗齿尖运动分析见图3.5所示,斗齿尖V点的坐标值XV和YV,是L1 、L2、L3的函数只要推导出XV和YV的函数表达式,那么整机作业范围就可以确定,现推导如下:由F点知:32= CFQ= 2 3 4 6 2 (3-20)在三角形CDF中:DCF由后面的设计确定,在DCF确定后则有: (3-21) (3-22) (3-23)在三角形DEF中 图3.5 齿尖坐标方程推导简图1则可以得斗杆瞬间转角2 (3-24)4、6在设计画图中确定。由三角形CFN知:l28 = Sqr(l162 + l12 - 2cos32l16l1) (3-25)由三角形CFQ知:l23 = Sqr(l22 + l12 - 2cos32l2l1) (3-26)由Q点知:35= CQV= 2 33 24 10 (3-27)在三角形CFQ中: (3-28)在三角形NHQ中: (3-29)在三角形HKQ中: (3-30)在四边形HNQK:NQH =24 + 26 (3-31)20 = KQV,其在后面的设计中确定。在列出以上的各线段的长度和角度之间的关系后,利用矢量坐标我们就可以得到各坐标点的值。(四) 特殊工作位置计算1、最大挖掘深度H1maxNH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖.图3.6 最大挖掘深度计算简图如图3.6示,当动臂全缩时,F, Q, V三点共线且处于垂直位置时,得最大挖掘深度为: H1max = YV = YFmin l2 l3 = YC + L1 Sin2 1min l2 l3 = YC + l1 Sin(1 20 11) l2 l3 (3-32)2、最大卸载高度H3maxNH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖图3.7 最大卸载高度计算简图如图3.7所示,当斗杆油缸全缩,动臂油缸全伸时,QV连线处于垂直状态时,得最大卸载高度为: (3-33)3、水平面最大挖掘半径R1maxNH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖图3.8 停机面最大挖掘半径计算简图如图3.8所示,当斗杆油缸全缩时,F、 Q、V三点共线,且斗齿尖v和铰点C在同一水平线上,即YC = YV,得到最大挖掘半径R1max为:R1max=XC+L40 (3-34)式中:L40 = Sqr(L1+L2+L3)2 2(L2+L3)L1COS32max (3-35)4、最大挖掘半径R2max NH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖图5.1 最大挖掘半径时工作装置结构简图最大挖掘半径时的工况是水平面最大挖掘半径工况下C、V连线绕C点转到水平面而成的。通过两者的几何关系,我们可计算得到:l 30 = 350mm ;l 40 = 5650mm。5、最大挖掘高度H2max最大挖掘高度工况是最大卸载高度工况中铲斗绕Q点旋转直到铲斗油缸全缩而形成的。具体分析方法和最大卸载高度工况的分析类似。四、 挖掘阻力分析(一)转斗挖掘阻力计算挖掘阻力可分为切向分力与法向分力,其中法向分力相对很小,一般为 (4-1) (4-2)在式(4-2)中,F1 切削阻力的切向分力;C土壤的硬度系数,对不同的土壤条件取值不同,这里设挖机用于级土壤的挖掘,取值为90;R铲斗与斗杆铰点到斗齿尖距离,即转斗切削半径其在前面已经初步确定,取值为90 cm;max挖掘过程中铲斗总转角的一半;现初定总转角为110,则max = 55某一挖掘位置处转斗的瞬时转角,B切削刃宽度影响系数,B = 1 + 2.6b = 1 + 2.60.8 = 3.08;A切削角变化影响系数,取A = 1.3.;Z带有斗齿的系数,取Z =0.75;X斗侧壁厚影响系数,X = 1+0.03S,其中S为侧壁厚度,单位为cm 。初步设计时取X = 1.15 ;D切削刃挤压土壤的力,根据经验统计和斗容量的大小选取D = 0.8 104N。当时,出现转斗挖掘最大切向分力,其值为: (4-3)将各参数代入式(4-3)得 转斗平均挖掘阻力按平均挖掘深度下的阻力计算,平均切削厚度为 (4-4)平均挖掘阻力为 (4-5) 将各参数代入上式得(二)斗杆挖掘阻力计算斗杆在挖掘过程中总转角一般为,现取。斗齿尖的行程实际上是斗杆转角所对应的弧长,根据经验公式有 (4-6)斗杆挖掘时切削半径,斗杆与动臂铰点至斗齿尖距离,单位m斗杆挖掘时切削厚度按如下公式计算 (4-7)q铲斗容量,B铲斗切削宽度m斗杆挖掘阻力计算公式如下: (4-8)式(4-8)中为挖掘阻力比,由附表010查得,对于级土取,对于,初步设计时取,将各参数代入式(4-8)得 取整为,斗杆挖掘阻力比转斗挖掘阻力要小一些,这是由于斗杆挖掘行程较长,切削厚度较小的缘故。五、基本尺寸的确定(一)斗形参数的确定斗容量q :在设计任务书中已给出q = 0.25 m3平均斗宽B:在设计任务书中已给出B = 0.8 m挖掘半径R:按经验统计和参考同斗容的其它型号的机械,初选R = 900mm 转斗挖掘装满转角(2):R、B及2三者与q之间有以几何关系q = 0.5 R2B(2-Sin2)KS在上式中:KS为土壤的松散系数,近似取值为1.25。将q = 0.25 m3和B = 0.8m代入上式有:铲斗两个铰点K、Q之间的间距l24和l3的比值k2的选取:l24太大将影响机构的传动特性,太小则影响铲斗的结构刚度3,一般取特性参数。初选特性参数k2 = 0.3。一般取。由于铲斗的转角较大,而k2的取值较小,故初选。(二)动臂机构参数的选择1、1与A点坐标的选取初选动臂弯角。由经验统计和参考其它同斗容机型,初选特性参数k3 = 1.65(k3 = L42/L41)铰点A坐标的选择:由底盘和转台结构,并结合同斗容其它机型的测绘,初选:XA = 560 mm ;YA = 700mm 2、 l1与l 2的选择 经统计分析,最大挖掘半径R1值与l1+l2+l3的值很接近,由已给定的最大挖掘半径R1、已初步选定的l3和k1,结合如下经验公式:; 式中: l1为动臂长, l 2为斗杆长,k1为动臂斗杆长度比将各参数代入上式得: ;3、 l41与l42的计算如图5.1所示,在三角形CZF中: NH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖图5.1 最大挖掘半径时工作装置结构简图l42 = k3l41 = 1.651407 = 2321 mm 4、l5的计算对于以反铲为主的通用挖掘机要适当考虑其他的换用装置(如正铲、起重等),而且要求在地面以上作业时能有足够的提升力矩,故初取k4 = 0.8511的取值对特性参数k4、最大挖掘深度H1max和最大挖高H2max均有影响,增大11会使k4减少或使H1max 增大,这符合反铲作业的要求,初选。斗杆液压油缸全缩时,CFQ =32 8最大,根据经验统计和便于计算,初选(32 8)max = 。由于采用单动臂液压缸,因此BCZ的取值较大,初取BCZ = 如上图5.1所示,在三角形CZF中:ZCF = 1 39 = - - = BCF = 2 =ZCF -ZCB由式(3-33)和式(3-34)有H3max = YC+ l1 Sin(1 20 11) l2 l3 (5-1) = YA+ l5 Sin11+ l1 Sin(1max 2 11)+ l2 Sin(1max+32 max 11 8 2 180) l3 H1max = l2 + l3 + l1 Sin(11 1min+ 2) l5 Sin11 YA ) (5-2)由式(5-1)、(5-2)有:H1max + H3max = l1 Sin(1max 2 11)+ l2 Sin(1max+ 32 max 11 8 2 180)+ l1 Sin(11 1min+ 2)+ l2 (5-3)令 A = 2+ 11 = + = B = A + (32 8)max = +()=将A、B的值代入式(5-3)中有H1max + H3max l1 Sin(1max ) Sin(1min ) + l2 Sin (1max +)1=0 又由特性参数 (5-4)则有 Sin1min = Sin1max 1 k4 = Sin1max1.36 (5-5) (5-6)将式(5-5)、式(5-6)代入到式(5-4)中得3500+3600-3400Sin(1max ) Sin(1min )+l2Sin(1max +)1 = 0 解之: 1max = ; 1min = 由式(5-2)有H1max = l2 + l3 + l1 Sin(11- 1min +2)- l5 Sin11- YA l5 = l2 + l3 + l1 Sin(11- 1min + 2)- YA - H1max Sin11 = 1700 + 900 + 3400Sin()- 800- 3500 Sin = 534.3mm1min与1max需要满足以下条件 (5-7) (5-8)将1max 、1min 的值代入式(5-7)、式(5-8)中得: = 0.482 = 1.316而 (5-9) (5-10)、满足5-9、5-10两个经验条件,说明、的取值是可行的。 (5-11) (5-12) (5-13) 至此,动臂机构的各主要基本参数已初步确定。(三) 动臂机构基本参数的校核1、动臂机构闭锁力的校核由第四章的计算可知,转斗的平均挖掘力由图5-2知,最大挖掘深度时的挖掘阻力力矩M1J:M1J = (H1max + YC) (5-14)式中,YC为C点的Y轴坐标值将各参数代入式(5-14)得 M1J = 0.312 105(3.5+1.162)= 1.45105 N.m 动臂油缸的闭锁力F1F1 = PgS1 (S1:动臂油缸小腔的作用面积) =2.1107(62.52 402)10 -6 = 1.5105 N 最大挖掘深度工作装置自身重力所产生的力矩MG :要求力矩,首先应该需要知道作用力和作用力臂。在此处,则是先要求出工作装置各部分的重量,由经验统计,初步估计工作装置的各部分重量如下:动臂G1 = 223kg 斗杆G2 = 179kg铲斗G3 = 86kg 斗杆缸G4 = 55kg铲斗缸G5 = 51kg 连杆机构G6 = 17kg动臂缸G7 = 55kg 图5.2 最大挖掘深度计算简图当处于最大挖掘深度时:1 = 1min = 由图5.2有MG (G1/2 +G2 +G3 +G4 +G5 +G6+ G7)10 l1 cos (5-15) =(111.5+179 +86 +55 +51 +17+55)103.4 cos = 1.5104N.m 动臂油缸的闭锁力与工作装置重力所产生的力矩(对C点的矩):M3 = F1l7 l5 Sin1min l1min + MG (5-16) = 21.51.459105 0.5343Sin40.51.109 + 1.5104 = 1.67105 N.m M1J = 1.45105 N.m 在式(5-16)中说明动臂油缸的闭锁力与工作装置重力所产生的力矩略大于平均挖掘阻力力矩,满足工作要求。2、满斗处于最大挖掘半径时动臂油缸提升力矩的校核 NH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖图5.3 最大挖掘半径时工作装置结构简图为方便计算,现将工作装置划分为二个部分,动臂、动臂液压缸和斗杆液压缸作为一部分,该部分重量以表示GB表示;其余的工作装置构件作为第二部分,重量以GG+D表示,于是有:GB=G1 +G4 +G7 =223 + 55 + 55 = 333kgGG+D =G2 +G3 +G5 +G6 = 179 + 86 + 51 + 17=333kg按经验公式取土的重量: GT = (1.6 1.8) q103 = 1.80.25 103 = 450kg当处于最大挖掘半径时,工作装置简图如图5.3所示,则有:MZ = 9.8GB l1 /2 + GG+D(l1 + 0.7l2)+ GT (l1 + l2 + l3 /2) = 9.83333.42+ 333(3.4+0.71.7)+ 450(3.4+1.7-0.92) = 0.45105 N.m 动臂油缸的推力: F1 = P1 S1 = 210762.5210- 6 = 2.45105 N在如图5.3所示,在三角形CAB中: (5-17)ACB =2 +11 +21 (5-18)将各参数分别代入式(5-17)和式(5-18)得 L1=1.542mL1 e1 = ACBCSinACB (5-19) 则此时动臂油缸提升力矩:MT = F1 e1= 2.451050.5054 =1.24105 N.m MZ = 0.45105 N.m 故铲斗处于最大挖掘半径时动臂油缸提升力矩满足工作要求。3、满斗处于最大高度时,动臂提升力矩的校核当斗杆在最大高度时的工况类似于图3.7,此时动臂油缸全伸,斗杆油缸全缩。1 =1max = 32 =32max = 2 = 21 = 1-(2 + 11) 37 = 32 -(- 21)则工作装置所受重力和土的重力所产生的载荷力矩MZ:MZ= (5-20)此时对于动臂油缸而言:L1 = L1max =1774 mm 1 =1max = 同式(5-19)的计算可求得此时的动臂油缸的力臂 此时动臂油缸的提升力矩MT可参考式(5-20)求得:MT = F1 e1 = 2010650210-60.388 = 0.61105 N.m MZ = 0.298105 N.m 说明满斗处于最大高度时,动臂提升力矩满足工作要求。E20(四)斗杆机构基本参数的选择E2ZD l92maxl8FD:斗杆油缸的下铰点;E:铲斗油缸的上铰点;F动臂的上铰点;2:斗杆的摆角;l9:斗杆油缸的最大作用力臂.图5.4斗杆机构基本参数计算简图取整个斗杆为研究对象,可得斗杆油缸最大作用力臂的表达式:e2max = l9 = F2d(l2 + l3 )/ P2 = 2104 (1700+900)10 -3/2010645210-6 = 409 mm (5-21) 如图5.4所示图中,D:斗杆油缸的下铰点;E:斗杆油缸的上铰点;F动臂的上铰点;2:斗杆的摆角;l9:斗杆油缸的最大作用力臂。斗杆油缸的初始位置力臂e20与最大力臂e2max有以下关系:e20 /e2max = l9 cos(2max /2)/l9 = cos (2max /2) (5-22)由5-22可知, 2max越大,则e20越小,即平均挖掘阻力越小.要得到较大的平均挖掘力,就要尽量减少2max,初取2max = 110由上图5.43的几何关系有:L2min = 2l9cSin (2max/2)/(2-1) = 2409Sin 55/(1.6 -1)= 1116.8 mm (5-23)L2max = L2min 2 = 1116.81.6= 1787 mm (5-24)l82 = L22min + l29 -2L2minl9cos( +2max)/2 = 1116.82+ 4092 + 21116.8409cos145 (5-25) l8 = 1470.6 mmEFQ取决于结构因素和工作范围,一般在130170之间,初定EFQ=160,动臂上DFZ也是结构尺寸,按结构因素分析,可初选DFZ=10。(五)铲斗机构基本参数的选择1、转角范围由最大挖掘高度H2max和最大卸载高度H3max的分析,可以得到初始转角:H2max-H3max = l3(Sin + 1.6) (5-26)将各参数代入式(5-26)得:5800-3600 = 900 (Sin + 1.6), = 53最大转角3max =V0QVZ,值太大会使斗齿平均挖掘力降低,常在150180之间选取,初选3max = 163。Kl292、铲斗机构其它基本参数的计算GL3Ml24l12FNQl21l2Vl3l12:摇臂的长度;l29:连杆的长度;l3:铲斗的长度;l2:斗杆的长度;F:斗杆的下铰点;G:铲斗油缸的下铰点;N:摇臂与斗杆的铰接点;K:铲斗的上铰点;Q:铲斗的下铰点.图5.5铲斗机构计算简图在图5.5中有:l24 = KQ = k2 l3 = 0.3900 = 270mmL3max 与L3min 的确定:由第四章的计算可知转斗平均挖掘阻力挖掘阻力F1P所做的W1p (5-27) 由图5-5,铲斗油缸推力所做的功W3:W3 = F3 (3-1)L3min = 2010650210-60.6L3min (5-28)由功的守恒知铲斗油缸推力所做的功W3 应该等于铲斗挖掘阻力所做的功W1p:即W3 = W1p (5-29)将5-27、5-28式代入5-29中计算可得:L3min = 849mm 圆整为850mm则L3max =3 L3min =1360mm剩余未选定的基本尺寸大部分为连杆机构尺寸,其应满足以下几个条件:1)挖掘力的要求:铲斗油缸的挖掘力应与转斗最大挖掘阻力相适应,当斗齿尖处于V1时,斗杆油缸的理论挖掘力应不低于最大挖掘阻力的80% 。 即PD080% PD0max;当处于最大理论挖掘力位置时V1QV应为30。2)几何相容。必须保证铲斗六连杆机构在l3全行程中任一瞬时都不会被破坏,即保证GFN、GHN以及四边形HNQK在任何瞬时皆成立。3)l3全行程中机构都不应出现死点,且传动角应当在允许的范围内。根据以上三个方面的要求,通过经验公式和同斗容的其它机型的测绘对照,初步选定剩余的基本尺寸如下:HK = 352mm; HN = 407mm;NQ = 300mm; FN = l2-NQ = 1400mm; GF =432mm;预选GFN = 60则 GN 2 = FN 2 + GF 2 2COSGFNFNGF GN = 1242mm至此,工作装置的基本尺寸均已初步确定。六、 工作装置结构设计整个工作装置由动臂、斗杆、铲斗及油缸和连杆机构组成,要确定这些构件的结构尺寸,必须要对其结构进行受力分析。要进行受力分析,首先要确定构件最不利的工况,并找到在该工况下的危险截面,以作为受力分析的依据。但构件在不利的工况下危险截面往往不止一个,这就需要分别计算出各危险截面尺寸再综合考虑,取其中的最大值作为最终的尺寸。(一)斗杆的结构设计1、斗杆的受力分析斗杆主要受到弯矩的作用,因此要找出斗杆中的最大弯矩进行设计计算。根据受力分析和以往的实验表明,在铲斗进行挖掘时,产生最大弯矩的工况满足以下条件:1)动臂处于最低位置。即动臂油缸全缩。2)斗杆油缸的力臂最大。3)铲斗齿尖在动臂与斗杆铰点和斗杆与铲斗铰点的连线上。4)侧齿挖掘时受到侧向力Wk的作用。在这个工况下斗杆会存在最大弯矩,受到的应力也会最大。该工况的具体简图如图6.1所示。取工作装置为研究对象,如图6.2所示。在该工况下存在的力有:工作装置各部件所受到的重力Gi;作用在铲斗上的挖掘阻力,包括切向阻力W1、法向阻力W2、侧向阻力W3。VNH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖图6.1 斗杆危险工况时的工作装置简图FNQPdW1HKW2G3HK-连杆 HN-摇臂N-摇臂与斗杆的铰接点 Q-斗杆与铲斗的铰接点图6.2 铲斗受力分析简图当动臂油缸全缩时,通过前面的章节可以得出21 = 45,由图6.1可知CF的向量可以表示为:FC = 3400COS(180-45)+Sin(180-45) = 3400(COS135+Sin135)由前面的章节计算结果知:ZFC =27,并初选DF = 1470mm。在DEF中DEF = 90COSEFD = EF/DF = 409/1470解得EFD = 73.8在CDEF中EFC = ZFC+DFZ+EFD = 27+10+73.8 = 110.8EFQ在前一章节已经初定为160由以上的角度关系知:FV = 2600Cos(360-110.8-160)+Sin(360-110.8-160) = 2600(Cos 89.2+Sin89.2) (6-1)OV = OC + CF + FV (6-2) = 1777(Cos87+Sin87)+3400(Cos-45+Sin-45)+ 2600(Cos 89.2+Sin89.2) 则XV = 1777Cos87 + 3400Cos(-45) + 2600 Cos(-89.2) = 1542 mm (6-3)由(3-16)式可i= 0.336则可得此时铲斗的理论挖掘力: F0D =F D i =1.651050.61 =1.0105 N切向阻力W1:初选该工况下铲斗重心到铰点Q的水平距离r2= l3 Cos(-89.2)/2=148mm取铲斗为研究对象,如图6.2所示,并对Q点取矩,则有MQ = 0(F0D - W1)l3 G3 r2 = 0(105- W1)0.9-8600.148 = 0W1 = 105 N法向阻力W2 的求解:工作装置所受重力对C点取矩有MC(Gi)= G1X1 +(G2 +G5)X2 + G3X3+G40.7XF+ G6X2 =2.231031.513 +(179+51)103.068+8602.837+5500.73.157 +1703.068 = 0.76105 N (6-4)W1到C点的距离r0r0 = l2 + l3CFCosCFV (6-5) = 1700+900-3400Cos(360-110.8-160) = 1481mm W2到C点的距离r1r1 = CFSinCFV = 3400Sin89.2 = 3210mm (6-6)法向阻力W2决定于动臂油缸的闭锁力F1 ,取整个工作装置为研究对象,则有MC = 0F1 e1+ MC(Gi )- W1 r0 - W2 r1 = 0 (6-7)将式(6-4)、(6-5)、(6-6)代入式(6-7)中解之得W2 = 0.32105 N 斗杆油缸作用力P2g的求解:FQ向量在X轴上的模值:XFN = FQ COS(-89.2) =17000.3291 = 560mm如图6.1所示,取斗杆(铲斗和连杆机构)为研究对象,则有:MC = 0P2gEF- W1 (l2+l3)- G3(XFN +r2)- G2XFN /2 = 0P2g0.41 -1052.6-860(0.56+0.148)-17900.56/2=0P2g=1.31105 N (6-8)而此时的斗杆闭锁力P2=21(45)2=1.34105 N,略大于P2g,说明闭锁力足够。横向挖掘阻力WK的求解:横向挖掘力WK由回转机构的制动器所承受,即WK的最大值决定于回转平台的制动力矩。故要先计算出制动力矩。地面附着力矩M:M = 6000G4/3 (其中 = 0.5) = 60000.5564/3 = 1.26105 N (6-9)在所设计的液压挖掘机中采用的是液压制动,由经验公式可求得回转机构的最大制动力矩MB:MB= 0.6M=0.756105 NWK = MB / XV = 0.756105/2.875 = 0.26105 N (6-10)Q点作用力与作用力矩RQx 、RQy、MQx、MQy的求解:取连杆机构为研究对象,如图6.3所示,则有:GRNP3NHX2KQY2RkNH-摇臂 HK-连杆G3-铲斗油缸的推力 RK连杆的作用力 RN摇臂的作用力图6.3 连杆机构计算简图X2 = 0PDCOSNHK-RNCOSHN X2 -RkCOSHKX2 = 0 (6-11)1.65105COS40.5-RNCOS57.5-RkCOS11.5 = 0 Y2 = 0PDSinGHK-RNSinHN X2-RkSinHKX2 = 0 (6-12)1.65105Sin40.5-RNSin57.5-RkSin11.5 = 0 由式(6-11)、(6-12)可解得:RN = -0.51105 N ; Rk =3.3105 N如图6.3所示,取整个铲斗为研究对象,以V点为新坐标的原点,VK为X3轴,过V点与VK垂直的直线为Y3,建立X3O3Y3坐标,则有:X3 = 0W2 -RQx -Rk COS11.5= 0 (6-13)0.32105-RQx3.3105 COS11.5 = 0RQx = -2.91105NY3 = 0RQy +W1- Rk Sin11.5= 0 (6-14)RQy +105 - 3.3105Sin11.5= 0RQy = -0.34105 NMQY3 = 0MQy -WK l3- W2 b/2= 0 (6-15)MQy -0.55 1051.55- 0.321050.52= 0 MQy = 105 NmMQX3 = 0MQxW1b/2= 0 (6-16) MQx = W1b/2=0.5 3105 NmN点作用力与作用力矩RNx 、RNy的求解:取曲柄和连杆为研究对象,如图6.4所示,则有:RNyF3HNRNxKRkH-摇臂 HK-连杆 F3-铲斗油缸的推力 RK连杆的作用力 RX摇臂的作用力沿HK连线上的分力 RY摇臂的作用力沿HK连线垂直方向上的分力图6.4 曲柄和连杆受力图X2 = 0RNX + Rk COS11.5- F3 = 0 RNX = 0.27105 NRNy = RNX tanFNH = 0.27105tan57.5=0.43105 N 2、斗杆内力图的绘制根据危险工况求出的斗杆所受到的力和力矩,可以绘制出在危险工况下的内力图,如图6.5、6.6、6.7、6.8、6.9、6.10所示。图6.5第一工况下斗杆的Nx图0.77105N0.54105N+0.34105NEFGNQ3.47105N图6.6 第一工况下斗杆的Qy图257KNm+FEQ图6.7 第一工况下斗杆的My图0.55105N+EFQ2.65105Nm图6.8 第一工况下斗杆的QZ图1105Nm+EFQ图6.9 第一工况下斗杆的MZ图0.53105Nm+FEQ图6.10 第一工况下斗杆的Tx图3、结构尺寸的计算由图6.8、图6.9、图6.10可知在通过F点且与斗杆下底板垂直的截面所受到的应力最大,是危险截面。故首先要对该截面进行计算,然后以此为基础再求解其它尺寸。 31斗杆宽度、钢板厚度、许用应力的选取由经验统计和其它同斗容机型的测绘,处取斗杆的宽度。挖掘机所用钢板的厚度在我国一般为,初选底板厚度如图6.11所示。12为斗杆侧板的厚度; 14为斗杆底板和顶板的厚度;238为底板的宽度图6.11在挖掘机中选用的结构钢材一般为16Mn,其有足够大的屈服极限和良好的机械性能。屈服极限。在斗杆中取安全系数,则斗杆的许用安全应力为:32斗杆危险截面处高度的计算危险截面的有效面积: (6-17) 该截面对y轴的惯性矩: (6-18) 该截面对z轴的惯性距: (6-19)横截面总面积: (6-20)该危险截面所受到的正应力: (6-21)该截面所受到的最大弯曲正应力: (6-22) (6-23)则截面所受到轴向拉应力与弯曲应力合成后有: (6-24)由于剪应力的大小相对于弯矩所产生的弯曲正应力要小得多,为简化计算,在计算中剪应力忽略不计,仅在校核中用,则有: (6-25)由式(6-21)、(6-22)、(6-23)、(6-24)、(6-25)解得h=360mm。有了危险截面的结构尺寸,再结合前面的基本尺寸,就可以利用CAD软件将斗杆绘制出来。这样斗杆的所有尺寸已经基本确定。(二)动臂结构设计同斗杆的受力分析及结构计算一样,首先还是要分析计算动臂可能出现最大应力的工况,找出在该工况下的危险截面并计算其尺寸。再以此为基础就可以计算出动臂上的其他尺寸。1、危险工况受力分析根据受力分析和以往的实验表明,在动臂上出现最大载荷的工况应满足以下条件:1)动臂油缸全缩。2)F、Q、V在同一条直线上,其连线与X轴垂直。3)铲斗挖掘时,斗边点遇到障碍。 该工况也就是最大挖掘深度工况,具体工作装置简图如6.12所示。NH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖图6.12 第一工况位置工作装置简图W1的求解:由于K、Q、V在同一条直线上,连杆机构的传动比不变,而铲斗的重力绕Q点所产生的力矩相对于铲斗油缸对C点所产生的力矩而言可以忽略不计,故W1的值与前面两工况一样,W1=105N。W2的求解:在此工况下时 ,而(前面的计算中已经得出)取整个工作装置为研究对象,则有: (6-26)求得为负值,在此工况中铲斗油缸的挖掘力不能得到最大的发挥。故需要转动铰点E直到铲斗油缸发挥最大挖掘力为止。由计算知当V点纵坐标即=-3000mm时,铲斗油缸能发挥最大的挖掘力。NH-摇臂;HK-连杆;C-动臂下铰点;A -动臂油缸下铰点;B-动臂与动臂油缸铰点;F-动臂上铰点;D-斗杆油缸上铰点;E-斗杆下铰点;G-铲斗油缸下铰点;Q-铲斗下铰点;K-铲斗上铰点;V-铲斗斗齿尖图6.13 实际工作时第一工况位置工作装置简图此工况是第工况下转动斗杆油缸而得的。除第2)点中的K、Q、V连线与X轴垂直修改成=-3000mm外,其他条件均不变,如图6.13所示。在此工况中,动臂油缸全缩,由前面的计算有: 则 解之 在DEF中,由几何关系则有:解得 ,而则由图6.17可知与的求解:由于挖掘时为铲斗油缸工作,而K、Q、V又在同一条直线上,故的值仍与前面的计算一样,。工作装置各部分受到的重力对C点的矩:(6-27)取整个工作装置为研究对象,则有: (6-28)也就是说此时仅是动臂与铲斗油缸进行挖掘。动臂铰点作用力的求取:取斗杆、铲斗、连杆机构为研究对象,则有: (6-29)方向与轴平行,在轴的正方向上。铰点的求解: (6-30) (6-31)对上下动臂附加弯矩与扭矩的求解:W与的夹角为,与的夹角为,则在坐标系上沿坐标轴的分力: (6-32) (6-33)则所产生的横向弯矩M: (6-34)则所产生的附加横向弯矩M: (6-35)所产生的附加横向扭矩T: (6-36)在坐标系上沿坐标轴的分力为: (6-37) (6-38)则所产生的附加横向弯矩: (6-39)所产生的附加扭矩: (6-40)2、内力图和弯矩图的求解上动臂所受到的轴向力: (6-41)上动臂所受到的剪力: (6-42)上动臂所受到的轴向弯矩: (6-43)下动臂所受到的轴向力: (6-44)下动臂所受到的剪力: (6-45)下动臂所受到的轴向弯矩: (6-46)由计算可以得出动臂在危险工况中的内力图和弯矩图如图6.14、6.15、6.16所示。C-动臂下铰点 F-动臂与动臂油缸铰点 Q-动臂油缸下铰点6.14 危险工况下N图C-动臂下铰点 F-动臂与动臂油缸铰点 Q-动臂油缸下铰点6.15危险工况下T图C-动臂下铰点 F-动臂与动臂油缸铰点 Q-动臂油缸下铰点6.16危险工况下M图3、结构尺寸计算由内力弯矩图分析知在动臂拐点处所受到的应力可能最大,是危险截面。因此我们首先要选择该截面进行计算,然后再以此为基础,就可以用作图法或计算得到动臂的其它结构尺寸。由现场测绘和经验统计,初步选择:动臂底板的宽度: 底板的厚度:由于上动臂所受的载荷较大,故取上动臂侧板的厚度,而下动臂所受的载荷相对要小,为方便制造与装配,选择下动臂的侧板的厚度也为。许用应力的选取:动臂钢板所选的材料为挖掘机中所普遍采用的低合金结构钢16Mn,其屈服极限,并初选安全系数则许用应力: (6-47)应力的计算与危险截面尺寸的求取:危险截面所围成的面积:危险截面所围成的有效面积: (6-48)则上动臂的有效面积: (6-49)下动臂的有效面积: (6-50)上动臂危险截面对Y轴的惯性矩: (6-51)下动臂危险截面对Y轴的惯性矩: (6-52)上动臂危险截面对Z轴的惯性矩: (6-53)同理下动臂危险截面对Z轴的惯性矩: (6-54)上动臂危险截面中:拉伸轴向力所产生的正应力: (6-55)弯曲所产生的正应力: (6-56)由应力的合成有: (6-57)解之:得到危险截面的尺寸后,利用作图法结合前面计算出来的尺寸就可以绘制出动臂图,从而也就得到了整个动臂的尺寸。(三)铲斗的设计1、铲斗斗形尺寸的设计反铲铲斗的斗形与尺寸,有较常用的经验统计公式,用户可以根据实际需要进行配制。根据经验公式和现场测绘,可以求得其中的未知参数。由经验公式初选:则下底板的斗形方程为:上顶板的斗形方程为:同理计算出铲斗抛物线部分的方程为:2、铲斗斗齿的结构计算铲斗的结构设计按最大弯矩进行设计,由力学分析知在与铲斗斗体连接处的弯矩最大,如图6.17所示,由公式6-58有: (6-58)a : 斗齿厚度 b : 斗齿宽度: 挖掘阻力 r : 斗齿尖到斗体的距离t : 铲斗的厚度 : 斗齿的许用应力代入值解得a=55 mma : 斗齿厚度 : 挖掘阻力r : 斗齿尖到斗体的距离 t : 铲斗的厚度6.17斗齿计算简图3、铲斗的绘制在铲斗的尺寸确定后,就可以用CAD软件进行绘制,绘制出的三维立体图的各视图如图6.18、6.19、6.20所示:6.18 铲斗三维视图的主视图6.19 铲斗三维视图的左视图6.20 铲斗三维视图的仰视图七、销轴与衬套的设计(一) 销轴的设计由于销轴与衬套的配合间隙较小,故以剪应力强度作为销轴的基本尺寸的设计,抗压强度与抗弯强度用于校核用。由有: (7-1)在设计计算时,应以所有工况中销轴所受到的剪应力最大值对销轴进行设计。在本设计中,销轴所选用的材料为30CrMnSi,其耐磨,在热处理后有着良好的综合机械性能。由于销轴在重载的较恶劣工况中工作,故选择。代入式7-1有:动臂各销轴的尺寸:斗杆各销轴的尺寸:(二)销轴用螺栓的设计螺栓选用的直径由销轴的直径不同分别选择两种系列的螺栓。(三) 衬套的设计为使衬套耐磨、减震与润滑性能好,选择衬套的材料为ZCuAl10Fe3。衬套的厚度选择为5mm,与销轴和圆筒分别采用间隙和过盈配合,如图7.1。则各销轴的尺寸为:7.1 衬套动臂各衬套的尺寸:斗杆各衬套的尺寸:八、总 结本次设计是在施工现场对液压小型挖掘机进行测绘的基础上,利用经验统计公式、旋转矢量法及力学计算,以机重为6t的挖掘机工作装置为对象,进行了以下设计工作:1、在现场测绘的基础上,结合经验公式进行了挖掘机工作装置的总体设计,并用旋转适量法对工作装置进行了运动学分析。2、用比例法和经验公式计算选择出工作装置各部分的基本尺寸。3、以上、下动臂、斗杆分别建立起三个新的坐标系,利用已经计算出的基本尺寸,对工作装置的各部分分别进行了力学分析。绘制出了工作装置斗杆和动臂的内力和弯矩图,选出了危险截面并计算出其结构尺寸。4、利用成熟的经验公式,选择出铲斗的斗形参数,并对铲斗斗齿进行力学分析以计算出其结构尺寸。5、对销轴和衬套进行了选材和尺寸的计算,并对其紧固的标准件进行了选型。挖掘机工作装置是挖掘机的核心部分,其结构的力学分析和计算十分复杂,难度也很大,设计不仅涉及到数学、力学、材料、公差、机械制造等方面的知识,还涉及到计算、绘图、文字分析表达等方面的综合实际能力。工作量之大,涉及面之广,都是前所未有的。而作者本身的知识面和能力都有限,不足之处还望各位老师、同学指正,以使设计不断完善。九、参考文献1 同济大学,太原重型机械学院.单斗液压挖掘机M.北京:中国建筑工业出版社,1980:40-86,264-274.2 金海薇. 液压挖掘机反铲工作装置CAD/CAM研究D.沈阳:辽宁工程技术大学,20013 刘本学.液压挖掘机反铲工作装置的有限元分析D.西安: 长安大学,2003,24 高衡、张全根主编.液压挖掘机 M.北京:中国建筑工业出版社,1981.8,74 -755 成大先主编.机械设计手册. 连接与紧固M.北京: 化学工业出版社,2004.16 胡传鼎编著. 机械制图画法范例M.北京: 化学工业出版社,2005.1,221-2277 杨晓辉主编. 简明机械实用手册 M.北京:科学出版社,2006.8,680-6898 范厚军主编.紧固件手册 M.南昌:江西科学出版社,2004.1,357-6409 机械设计手册编委会.机械设计手册第一卷 M.北京:机械工业出版社,2004.8,10、徐灏主编.机械设计手册(4).北京:机械工业出版社,199111、王文斌.机械设计手册(3).北京:机械工业出版社,200412、孟少龙主编.机械加工工艺手册.北京:机械工业出版社,199113、中国出版社第三编辑室主编.公差与配合标准手册.北京:中国标准出版社,199614、章宏甲、黄谊主编.液压传动.北京:机械工业出版社,200215、刘希平主编.工程机械构造图册. 北京:机械工业出版社,199016、陈育仪编著.工程机械优化设计.北京:中国铁道出版社,198717、孔德文,赵克利,徐宁生主编. 液压挖掘机,化学工业出版社, 2007,1018、唐增宝,常建娥主编. 机械设计课程设计(第3版),华中科技大学出版社, 200619、张铁主编. 液压挖掘机结构、原理及使用,石油大学出版社,2002,1220 Nease, A.D.; Alexander, E.F, Air Force construction automation/robotics. In Proc. 10th International Symposium on Automation and Robotics in Construction (ISARC); Houston, May 1993. 十、致 谢在本次设计完成之际,首先要感谢我的指导老师刘柏希教授,刘教授治学严谨,工作仔细认真,教给了我许多先进的设计方法和设计理念,在我困惑时耐心地给予分析和讲解,正是在她的大力帮助和支持下,才使这次设计得以顺利完成。也让我在学习新事物、检索资料、绘图、分析并解决问题等各方面的能力得到了较大的锻炼和提高,巩固和加深了所学的知识,培养了独立工作的能力。其次要感谢与我同做一个课题的二位同学,他们在我的毕业设计过程中给了我很大的帮助和支持。感谢机械学院所有无私传授知识的各位老师。感谢在百忙中抽出时间来评阅本文的各位老师。附录附录1Design of machine and machine elementsMachine designMachine design is the art of planning or devising new or improved machines to accomplish specific purposes. In general, a machine will consist of a combination of several different mechanical elements properly designed and arranged to work together, as a whole. During the initial planning of a machine, fundamental decisions must be made concerning loading, type of kinematic elements to be used, and correct utilization of the properties of engineering materials. Economic considerations are usually of prime importance when the design of new machinery is undertaken. In general, the lowest over-all costs are designed. Consideration should be given not only to the cost of design, manufacture the necessary safety features and be of pleasing external appearance. The objective is to produce a machine which is not only sufficiently rugged to function properly for a reasonable life, but is at the same time cheap enough to be economically feasible. The engineer in charge of the design of a machine should not only have adequate technical training, but must be a man of sound judgment and wide experience, qualities which are usually acquired only after considerable time has been spent in actual professional work.Design of machine elements The principles of design are, of course, universal. The same theory or equations may be applied to a very small part, as in an instrument, or, to a larger but similar part used in a piece of heavy equipment. In no ease, however, should mathematical calculations be looked upon as absolute and final. They are all subject to the accuracy of the various assumptions, which must necessarily be made in engineering work. Sometimes only a portion of the total number of parts in a machine are designed on the basis of analytic calculations. The form and size of the remaining parts are designed on the basis of analytic calculations. On the other hand, if the machine is very expensive, or if weight is a factor, as in airplanes, design computations may then be made for almost all the parts. The purpose of the design calculations is, of course, to attempt to predict the stress or deformation in the part in order that it may sagely carry the loads, which will be imposed on it, and that it may last for the expected life of the machine. All calculations are, of course, dependent on the physical properties of the construction materials as determined by laboratory tests. A rational method of design attempts to take the results of relatively simple and fundamental tests such as tension, compression, torsion, and fatigue and apply them to all the complicated and involved situations encountered in present-day machinery. In addition, it has been amply proved that such details as surface condition, fillets, notches, manufacturing tolerances, and heat treatment have a market effect on the strength and useful life of a machine part. The design and drafting departments must specify completely all such particulars, must specify completely all such particulars, and thus exercise the necessary close control over the finished product. As mentioned above, machine design is a vast field of engineering technology. As such, it begins with the conception of an idea and follows through the various phases of design analysis, manufacturing, marketing and consumerism. The following is a list of the major areas of consideration in the general field of machine design: Initial design conception; Strength analysis; Materials selection; Appearance; Manufacturing; Safety; Environment effects; Reliability and life; Strength is a measure of the ability to resist, without fails, forces which cause stresses and strains. The forces may be; Gradually applied; Suddenly applied; Applied under impact; Applied with continuous direction reversals; Applied at low or elevated temperatures. If a critical part of a machine fails, the whole machine must be shut down until a repair is made. Thus, when designing a new machine, it is extremely important that critical parts be made strong enough to prevent failure. The designer should determine as precisely as possible the nature, magnitude, direction and point of application of all forces. Machine design is mot, however, an exact science and it is, therefore, rarely possible to determine exactly all the applied forces. In addition, different samples of a specified material will exhibit somewhat different abilities to resist loads, temperatures and other environment conditions. In spite of this, design calculations based on appropriate assumptions are invaluable in the proper design of machine. Moreover, it is absolutely essential that a design engineer knows how and why parts fail so that reliable machines which require minimum maintenance can be designed. Sometimes, a failure can be serious, such as when a tire blows out on an automobile traveling at high speeds. On the other hand, a failure may be no more than a nuisance. An example is the loosening of the radiator hose in the automobile cooling system. The consequence of this latter failure is usually the loss of some radiator coolant, a condition which is readily detected and corrected. The type of load a part absorbs is just as significant as the magnitude. Generally speaking, dynamic loads with direction reversals cause greater difficulties than static loads and, therefore, fatigue strength must be considered. Another concern is whether the material is ductile or brittle. For example, brittle materials are considered to be unacceptable where fatigue is involved. In general, the design engineer must consider all possible modes of failure, which include the following: Stress; Deformation; Wear; Corrosion; Vibration; Environmental damage; Loosening of fastening devices. The part sizes and shapes selected must also take into account many dimensional factors which produce external load effects such as geometric discontinuities, residual stresses due to forming of desired contours, and the application of interference fit joint. Selected from” design of machine elements”, 6th edition, m. f. sports, prentice-hall, inc., 1985 and “machine design”, Anthony Esposito, charles e., Merrill publishing company, 1975.Mechanical properties of materials The material properties can be classified into three major headings: (1) physical, (2) chemical, (3) mechanicalPhysical properties Density or specific gravity, moisture content, etc., can be classified under this category. Chemical propertiesMany chemical properties come under this category. These include acidity or alkalinity, react6ivity and corrosion. The most important of these is corrosion which can be explained in laymans terms as the resistance of the material to decay while in continuous use in a particular atmosphere. Mechanical properties Mechanical properties include in the strength properties like tensile, compression, shear, torsion, impact, fatigue and creep. The tensile strength of a material is obtained by dividing the maximum load, which the specimen bears by the area of cross-section of the specimen. This is a curve plotted between the stress along the This is a curve plotted between the stress along the Y-axis(ordinate) and the strain along the X-axis (abscissa) in a tensile test. A material tends to change or changes its dimensions when it is loaded, depending upon the magnitude of the load. When the load is removed it can be seen that the deformation disappears. For many materials this occurs op to a certain value of the stress called the elastic limit Ap. This is depicted by the straight line relationship and a small deviation thereafter, in the stress-strain curve (fig.3.1). Within the elastic range, the limiting value of the stress up to which the stress and strain are proportional, is called the limit of proportionality Ap. In this region, the metal obeys hookess law, which states that the stress is proportional to strain in the elastic range of loading, (the material completely regains its original dimensions after the load is removed). In the actual plotting of the curve, the proportionality limit is obtained at a slightly lower value of the load than the elastic limit. This may be attributed to the time-lagin the regaining of the original dimensions of the material. This effect is very frequently noticed in some non-ferrous metals. Which iron and nickel exhibit clear ranges of elasticity, copper, zinc, tin, are found to be imperfectly elastic even at relatively low values low values of stresses. Actually the elastic limit is distinguishable from the proportionality limit more clearly depending upon the sensitivity of the measuring instrument. When the load is increased beyond the elastic limit, plastic deformation starts. Simultaneously the specimen gets work-hardened. A point is reached when the deformation starts to occur more rapidly than the increasing load. This point is called they yield point Q. the metal which was resisting the load till then, starts to deform somewhat rapidly, i. e., yield. The yield stress is called yield limit Ay. The elongation of the specimen continues from Q to S and then to T. The stress-strain relation in this plastic flow period is indicated by the portion QRST of the curve. At the specimen breaks, and this load is called the breaking load. The value of the maximum load S divided by the original cross-sectional area of the specimen is referred to as the ultimate tensile strength of the metal or simply the tensile strength Au. Logically speaking, once the elastic limit is exceeded, the metal should start to yield, and finally break, without any increase in the value of stress. But the curve records an increased stress even after the elastic limit is exceeded. Two reasons can be given for this behavior: The strain hardening of the material; The diminishing cross-sectional area of the specimen, suffered on account of the plastic deformation. The more plastic deformation the metal undergoes, the harder it becomes, due to work-hardening. The more the metal gets elongated the more its diameter (and hence, cross-sectional area) is decreased. This continues until the point S is reached. After S, the rate at which the reduction in area takes place, exceeds the rate at which the stress increases. Strain becomes so high that the reduction in area begins to produce a localized effect at some point. This is called necking. Reduction in cross-sectional area takes place very rapidly; so rapidly that the load value actually drops. This is indicated by ST. failure occurs at this point T. Then percentage elongation A and reduction in reduction in area W indicate the ductility or plasticity of the material: A=(L-L0)/L0*100% W=(A0-A)/A0*100% Where L0 and L are the original and the final length of the specimen; A0 and A are the original and the final cross-section area. Selected from “testing of metallic materials”Quality assurance and control Product quality is of paramount importance in manufacturing. If quality is allowed deteriorate, then a manufacturer will soon find sales dropping off followed by a possible business failure. Customers expect quality in the products they buy, and if a manufacturer expects to establish and maintain a name in the business, quality control and assurance functions must be established and maintained before, throughout, and after the production process. Generally speaking, quality assurance encompasses all activities aimed at maintaining quality, including quality control. Quality assurance can be divided into three major areas. These include the following:Source and receiving inspection before manufacturing;In-process quality control during manufacturing;Quality assurance after manufacturing. Quality control after manufacture includes warranties and product service extended to the users of the product. Source and receiving inspection before manufacturing Quality assurance often begins ling before any actual manufacturing takes place. This may be done through source inspections conducted at the plants that supply materials, discrete parts, or subassemblies to manufacturer. The manufacturers source inspector travels to the supplier factory and inspects raw material or premanufactured parts and assemblies. Source inspections present an opportunity for the manufacturer to sort out and reject raw materials or parts before they are shipped to the manufacturers production facility. The responsibility of the source inspector is to check materials and parts against design specifications and to reject the item if specifications are not met. Source inspections may include many of the same inspections that will be used during production. Included in these are:Visual inspection;Metallurgical testing;Dimensional inspection;Destructive and nondestructive inspection;Performance inspection.Visual inspections Visual inspections examine a product or material for such specifications as color, texture, surface finish, or overall appearance of an assembly to determine if there are any obvious deletions of major parts or hardware. Metallurgical testing Metallurgical testing is often an important part of source inspection, especially if the primary raw material for manufacturing is stock metal such as bar stock or structural materials. Metals testing can involve all the major types of inspections including visual, chemical, spectrographic, and mechanical, which include hardness, tensile, shear, compression, and spectr5ographic analysis for alloy content. Metallurgical testing can be either destructive or nondestructive. Dimensional inspection Few areas of quality control are as important in manufactured products as dimensional requirements. Dimensions are as important in source inspection as they are in the manufacturing process. This is especially critical if the source supplies parts for an assembly. Dimensions are inspected at the source factory using standard measuring tools plus special fit, form, and function gages that may required. Meeting dimensional specifications is critical to interchangeability of manufactured parts and to the successful assembly of many parts into complex assemblies such as autos, ships, aircraft, and other multipart products. Destructive and nondestructive inspection In some cases it may be necessary for the source inspections to call for destructive or nondestructive tests on raw materials or p0arts and assemblies. This is particularly true when large amounts of stock raw materials are involved. For example it may be necessary to inspect castings for flaws by radiographic, magnetic particle, or dye penetrant techniques before they are shipped to the manufacturer for final machining. Specifications calling for burn-in time for electronics or endurance run tests for mechanical components are further examples of nondestructive tests. It is sometimes necessary to test material and parts to destruction, but because of the costs and time involved destructive testing is avoided whenever possible. Examples include pressure tests to determine if safety factors are adequate in the design. Destructive tests are probably more frequent in the testing of prototype designs than in routine inspection of raw material or parts. Once design specifications are known to be met in regard to the strength of materials, it is often not necessary to test further parts to destruction unless they are genuinely suspect. Performance inspection Performance inspections involve checking the function of assemblies, especially those of complex mechanical systems, prior to installation in other products. Examples include electronic equipment subcomponents, aircraft and auto engines, pumps, valves, and other mechanical systems requiring performance evaluation prior to their shipment and final installation. Selected form “modern materials and manufacturing process”Electro-hydraulic drum brakesApplication The YWW series electro-hydraulic brake is a normally closed brake, suitable for horizontal mounting. It is mainly used in portal cranes, bucket stacker/reclaimersslewing mechanism.The YKW series electro-hydraulic brake is a normally opened brake, suitable for horizontal mounting, employing a thruster as actuator. with the foot controlling switch the operator can release or close the brake. It is mainly used for deceleration braking of portal cranesslewing mechanism. In a non-operating state the machinery can be braked by a manual close device.The RKW series brake is a normally opened brake, which is operated by foot driven hydraulic pump, suitable for horizontal mounting. Mainly used in the slewing mechanism of middle and small portal cranes. When needed, the brake is activated by a manual closed device. Main design featuresInterlocking shoes balancing devices (patented technology) constantly equalizes the clearance of brake shoes on both sides and made adjustment unnecessary, thus avoiding one side of the brake lining sticking to the brake wheel. The brake is equipped with a shoed autoaligning device.Main hinge points are equipped with self-lubricating bearing, making high efficiency of transmission, long service life. Lubricating is unnecessary during operation.Adjustable bracket ensure the brake works well.The brake spring is arranged inside a square tube and a surveyors rod is placed on one side. It is easy to read braking torque value and avoid measuring and computing.Brake lining is of card whole-piece shaping structure, easy to replace. Brake linings of various materials such as half-metal (non-asbestos) hard and half-hard, soft (including asbestos) substance are available for customers to choose.All adopt the companys new types of thruster as corollary equipment which work accurately and have long life. Hydraulic Power TransmissionThe Two Types Of Power Transmission In hydraulic power transmission the apparatus (pump) used for conversion of the mechanical (or electrical,thermal) energy to hydraulic energy is arranged on the input of the kinematic chain ,and the apparatus (motor) used for conversion of the hydraulic energy to mechanical energy is arranged on the output (fig.2-1) The theoretical design of the energy converters depends on the component of the bernouilli equation to be used for hydraulic power transmission. In systerms where, mainly, hydrostatic pressure is utilized, displacement (hydrostatic) pumps and motors are used, while in those where the hydrodynamic pressure is utilized is utilized gor power transmission hydrodynamic energy converters (e.g. centrifugal pumps) are used. The specific characteristic of the energy converters is the weight required for transmission of unit power. It can be demonstrated that the use of hydrostatic energy converters for the low and medium powers, and of hydrodynamic energy converters of high power are more favorite (fig.2-2). This is the main reason why hydrostatic energy converters are used in industrial apparatus. transformation of the energy in hydraulic transmission. 1.driving motor (electric, diesel engine);2.mechanical energy;3.pump; 4.hydraulic energy; 5.hydraulic motor; 5.mechanical energy; 6.load variation of the mass per unit power in hydrostatic and hydrodynamic energy converters 1、hydrostatic; 2.hydrodynamicOnly displacement energy converters are dealt with in the following. The elements performing converters provide one or several size. Expansion of the working chambers in a pump is produced by the external energy admitted, and in the motor by the hydraulic energy. Inflow of the fluid occurs during expansion of the working chamber, while the outflow (displacement) is realized during contraction. Such devices are usually called displacement energy converters. The Hydrostatic Power In order to have a fluid of volume V1 flowing in a vessel at pressure work spent on compression W1 and transfer of the process, let us imagine a piston mechanism (fig.2-3(a) which may be connected with the aid of valves Z0 and Z1 to the external medium under pressure P0 and reservoir of pressure p1.in the upper position of the piston (x=x0) with Z0 open the cylinder chamber is filled with fluid of volume V0 and pressure P0. now shut the value Z0 and start the piston moving downwards. If Z1 is shut the fluid volume in position X=X1 of the piston decreases from V0 to V1, while the pressure rises to P1. the external work required for actuation of the piston (assuming isothermal change) is W1=-0x0(P-P0)Adx=-v1v0(P-P0)dvSelect from Hydraulic Power Transmission机器和机器零件的设计机器设计机器设计为了特定的目的而发明或改进机器的一种艺术。一般来讲,机器时有多种不同的合理设计并有序装配在一起的部件构成的,在最初的机器设计阶段,必须基本明确负载、元件的运动情况、工程材料的合理使用性能。负责新机器的设计最初的最重要的是经济性考虑。一般来说,选择总成本最低的设计方案,不仅要考虑设计、制造、销售、安装的成本。还要考虑服务的费用,机械要保证必要的安全性能和美观的外形。制造机器的目标不仅要追求保证只用功能的合理寿命,还要保证足够便宜以同时保证其经济的可行性。负责设计机器的工程师,不仅要经过专业的培训,而且必须是一个准确判断而又有丰富经验的人,具有一种有足够时间从事专门的实际工作的素质。机器零件的设计相同的理论或方程可应用在一个一起的非常小的零件上,也可用在一个复杂的设备的大型相似件上,既然如此,毫无疑问,数学计算是绝对的和最终的。他们都符合不同的设想,这必须由工程量决定。有时,一台机器的零件全部计算仅仅是设计的一部分。零件的结构和尺寸通常根据实际考虑。另一方面,如果机器和昂贵,或者质量很重要,例如飞机,那麽每一个零件都要设计计算。当然,设计计算的目的是试图预测零件的应力和变形,以保证其安全的带动负载,这是必要的,并且其也许影响到机器的最终寿命。当然,所有的计算依赖于这些结构材料通过试验测定的物理性能。国际上的设计方法试图通过从一些相对简单的而基本的实验中得到一些结果,这些试验,例如结构复杂的及现代机械设计到的电压、转矩和疲劳强度。另外,可以充分证明,一些细节,如表面粗糙度、圆角、开槽、制造公差和热处理都对机械零件的强度及使用寿命有影响。设计和构建布局要完全详细地说明每一个细节,并且对最终产品进行必要的测试。综上所述,机械设计是一个非常宽的工程技术领域。例如,从设计理念到设计分析的每一个阶段,制造,市场,销售。以下是机械设计的一般领域应考虑的主要方面的清单:最初的设计理念 受力分析 材料的选择 外形 制造 安全性 环境影响 可靠性及寿命在没有破坏的情况下,强度是抵抗引起应力和应变的一种量度。这些力可能是:渐变力 瞬时力 冲击力 不断变化的力 温差如果一个机器的关键件损坏,整个机器必须关闭,直到修理好为止。设计一台新机器时,关键件具有足够的抵抗破坏的能
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