JWB-75型无极绳绞车设计【含6张CAD图纸、说明书】
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含6张CAD图纸、说明书
无极绳绞车设计【
张CAD图纸
设计含CAD图纸
CAD图纸设计
型无极绳绞车
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摘 要JWB-75型无极绳绞车是一种新型绞车，主要用于煤矿井下工作面顺槽和采区大巷材料、设备、液压支架等辅助运输，也可用于金属矿井下巷道和地面轨道运输，适用于坡度不大有起伏变化的轨道运输，最大起伏度16，适用运输距离约为2000m. JWB-75型无极绳绞车主要由电机、联轴器、制动器、变速箱、卷筒装置、底座等部分组成。在传动系统方面它采用了一个闭环传动系统：圆锥齿轮传动斜齿圆柱齿轮传动行星齿轮传动。其传动技术原理简单、可靠、高效、节能。 该绞车具有良好的防爆性和制动性能，并且具有使用寿命长，传动效率高，结构紧凑的特点。绞车采用抛物线滚筒，能够实现正反向工作。绞车整体呈L型结构，底座呈雪橇状，在井下自移平稳。绞车中心低，底座刚性好，而且安装方便，安全可靠。关键词：无极绳绞车； 最大起伏度； 闭环传动； 正反向工作ABSTRACTThe type of JWB-75 endless rope winch is a new type. It can be used in conveying materials, equipments and hydraulic supports in working area roadway and working surface in the underground of coal mines. Besides it can be used in the metal mines to convey in the roadway and on the ground track. The winch applicable to the track transportation with little slope .Maximum amplitude don not exceed 16,the appropriate transportation distance is about 2000m.The model of JWB-75 endless rope winch consists of A.C.motor, clutch, electrical-hydraulic brake, transmission, roller, pedestal and so on. In the aspect of transmission system, it realizes a closed transmitting circuit which has three step: from the taper gear to the cylinder gear to the planet gear. Its transmission technology principle is simple, reliable, high efficiency and energy saving. And this also makes the winch have the ability of big tonnage and brake-permitting. This winch is of explosion-proof type. It has good braking performance, long enwind rope, long enterance and high clriving effectiveness. It has compact construction, The winch has a parabolic-type drum so as to achieve positive and negative work. The pedestal is of sledge form. So it can be move in the underground shaft. The gravity center of the winch is low and the rigidity of pedstal is good, which make the winch have the features of stable operation, expedient installation, safe and reliable.Keywords: endless rope winch; maximum amplitude; closed transmitting circuit; positive and negative work目 录1 绪论11.1背景11.2 绞车简介11.2.1对无极绳的概述21.2.2无极绳绞车21.3 设计目的32 设计方案32.1 JWB无极绳绞车的组成32.2绞车的工作原理42.3主要构成及结构特征43 总体设计53.1 设计总则53.2 已知条件53.3 牵引钢丝绳直径及滚筒直径的确定53.3.1钢丝绳的选择53.3.2 滚筒参数的确定53.4 电机的选择、传动系统的确定和运动学计算63.4.1 电动机选型63.4.2 传动系统的确定73.4.3计算传动效率74 齿轮传动设计计算84.1传动装置运动参数的计算84.1.1 各轴转速计算84.1.2 各轴功率计算84.1.3 各轴扭矩计算84.2锥齿轮传动设计计算84.2.1选材料、热处理方法，定精度等级。84.2.2初步设计84.2.3 几何尺寸计算94.2.4 校核接触强度134.2.5齿根弯曲强度校核144.2.6 齿轮各检验项目及公差值的计算164.3斜齿圆柱齿轮传动设计计算174.3.1选择材料，确定和以及精度等级174.3.2 按接触强度进行初步设计174.3.3 校核齿面接触强度204.3.4 校核齿根弯曲强度214.3.5 齿轮及齿轮副精度精度的检验项目计算234.4 行星齿轮传动设计244.4.1 选取行星传动齿轮传动的传动类型和传动简图244.4.2 配齿计算244.4.3材料选择及热处理方式244.4.4a-c 齿轮按接触强度初步计算254.4.5 a c齿轮传动的主要尺寸254.4.6 b-c齿轮传动的主要尺寸264.4.7内齿圈B主要尺寸274.4.8 装配条件的验算274.4.9 强度验算275 轴的设计345.1 小锥齿轮轴的设计计算与强度校核345.1.1 轴的初步计算345.1.2 轴的疲劳强度校核345.1.3 锥齿轮轴上轴承的寿命验算375.2 大锥齿轮轴的设计计算和强度验算375.2.1 选材料375.2.2 估算轴的直径375.2.3 结构设计385.2.4 齿轮上的作用力385.2.5轴上键的强度验算425.2.6轴上轴承寿命校核425.3 大斜齿圆柱齿轮轴的设计计算与强度校核435.3.1估算轴的直径435.3.2 轴的疲劳强度校核435.3.3 轴上键的强度校核465.3.4 轴承寿命的验算485.4 行星轮轴的设计计算与强度校核485.4.1 初定轴的直径485.4.2 轴的疲劳强度校核485.4.3 轴承的寿命校核495.5 行星架轴的设计计算与强度校核505.5.1估算轴的直径505.5.2 渐开线花键的强度校核505.6 卷筒轴的设计525.6.1估算轴的直径525.6.2 轴的疲劳强度校核525.6.3 轴承的寿命校核535.6.4 键的校核.545.7 轴的受力简图546 JWB-75型无极绳绞车使用说明书5661使用范围5662主要技术规范5663绞车的润滑与密封5764绞车的装配、调整、及试运转576.4.1变速箱576.4.2 卷筒装置576.4.3总装586.4.4 空负荷试运转586.4.5 负荷试运转586.5 绞车的固定方法和安设位置586.5.1 绞车在现场的固定方法586.6 绞车的操作规程596.6.1 工作前的注意事项596.7绞车的维护、拆卸与修理597 结论64参考文献65翻译部分66英文原文66中文翻译72致 谢761 绪论1.1背景建国以来，我国地方煤炭的发展非常迅速，产量成倍增长。当前地方煤矿已成为煤炭工业的一个重要阵地，对国民经济的全面发展，改善煤炭工业布局，发展地方经济，支援农业,解决城乡人民生活需要方面，起了很好的作用。煤炭是当前我国能源的重要组成部分之一，是我国实现工业现代化的重要能源基础。我国的煤炭工业发展还不能适应整个国民经济发展的需要，为了适应整个国民经济持续、快速、高效发展的需要，必须以更快的速度发展煤炭工业。然而，高速发展煤炭工业的出路在于机械化。近40年我国的煤炭工业发生了巨大变化，开采技术已接近或部分达到国际先进水平，综采单采原煤产量早已突破了百万吨，然而煤炭工业机械化离不了运输，运输又离不了辅助运输设备，绞车就是辅助运输设备的一种。原煤的运输也已经实现了大运量带式输送机化，但井下轨道辅助运输与之很不适应，材料的运输基本上沿用传统的小绞车群接力式的运输方式，运输战线长，环节多，占用的搬运设备、人员多，安全性差，效率低。尽管一些煤矿对其进行了技术改造, 但仍然满足不了当前矿井发展和生产的需要。可见矿井辅助运输是当前现代化矿井建设的关键和重点。我国绞车的诞生是从20世纪50年代开始的，初期主要仿制日本和苏联的；60年代进入了自行设计阶段；到了70年代，随着技术的慢慢成熟，绞车的设计也进入了标准化和系列化的发展阶段。但与国外水平相比，我国的绞车在品种、型式、结构、产品性能，三化水平（参数化、标准化、通用化）和技术经济方面还存在一定的差距。国外矿用绞车发展趋势有以下几个特点：a、标准化系列化；b、体积小、重量轻、结构紧凑；c、高效节能；d、寿命长、低噪音；e、一机多能、通用化；f、大功率；g、外形简单、平滑、美观、大方。针对国外的情况我们应该采取以下措施：a、制定完善标准，进行产品更新改造和提高产品性能；b、完善测试手段，重点放在产品性能检测；c、技术引进和更新换代相结合；d、组织专业化生产，争取在较短时间内达到先进国家的水平。1.2 绞车简介现有的绞车类型有：（1）JTK系列提升绞车。主要用于煤矿及其他矿山作为斜坡的提升设备，或作为其它辅助拖运工具。有防爆型和非防爆型两种。防爆型配有防爆电机及电器，可用于煤矿井下有煤尘和瓦斯的地区。不能用于升降人员。（2）JDM系列调车绞车。适用于煤矿、选煤厂、电厂、钢铁厂、钢金矿场、水泥厂、港口等场所调度列车之用。JDM型为单摩擦轮型式，适用于单向调车。2JDM型为双摩擦轮型式，适用于双向调车。其由电动机、减速器、摩擦轮组成，具有结构先进、体积小、重量轻、效率高、操作方便、投资费用少、维护费用低等特点。（3）JHS型双速系列回柱绞车。均具有隔爆性能，主要适用于煤矿井下回柱放顶，亦可在运输、牵引中作运料、移溜槽等辅助工作。在结构上采用了球面蜗轮副传动，手动拨叉实现换挡调速，快速可到慢速的56倍，可大提高工作效率，与传统的双速纺车相比有结构紧凑、体积小、噪音低、运转平稳、调速灵活、安全可靠、手动刹车、断电自锁等特点。（4）JY系列绞车均为隔爆型，供矿井井下或地面装载调度编组矿车及在中间巷拖运矿车和其它辅助搬运工作。JY系列绞车均采用行星齿轮传动，绞车具有结构紧凑、刚性好、效率高、安装移动方便，起动平稳、操作灵活、制动可靠、噪音低等特点。1.2.1 对于无极绳的概述无极绳运输是矿山辅助运输的重要组成部分，是利用钢丝绳在水平或者倾斜的轨道上牵引矿车或提升容器沿轨道移动的一种运输方式。目前，在我国当前的矿山企业中，特别是中小型煤矿的地面运输以及主要运输巷道、中间平巷、采区上、下山运输等地方都可以考虑采用无极绳运输。按照钢丝绳与矿车相对位置不同，无极绳运输分为上绳式（钢丝绳在矿车的上方）和下绳式（钢丝绳在矿车的下方）两种绳式（如图1-1）。上绳式运行阻力小，但在线路上需要铺设许多托绳轮和支架，消耗材料多。由于钢丝绳搁在矿车上，两者磨损都很大；在起伏的底板上，矿车运行欠稳定。因此，上绳式已淘汰。现在主要采用下绳式，下绳式的优缺点与上述相反。 （a）上绳式 （b）下绳式图1-1 无极绳运输类型1.钢丝绳 2.矿车 3.轨道无极绳运输与单滚筒、双滚筒等有极绳运输相比，具有以下优点:（1）牵引力恒定平稳、爬坡能力强；（2）结构简单、可靠性高、维护量小；（3）可实现连续运输、运输效率高；（4）对巷道适应性强、使用方便；（5）操作简单、安全；（6）无污染及噪音。但无极绳运输却有如下缺点：（1）只能实现固定巷道运输、不具有机动性；（2）绞车司机需依靠信号操作、直观性差；（3）在挂车、摘车时需其它设备调车；（4）沿线布置有钢丝绳，需对沿线道岔作处理。1.2.1 无极绳绞车无极绳绞车是为无极绳运输提供动力的设备，是摩擦式绞车，由电动机、减速器、摩擦滚筒、制动装置和操纵机构等组成，除了摩擦滚筒外与一般的绞车相同。目前国内有很多厂家生产制造无极绳绞车，主要有两种结构形式（如图1-2）。 （a）右型 （b）左型图1-2 无极绳绞车 1.摩擦滚筒 2.减速器 3.电动机 国产无极绳绞车分为螺旋缠绕式和夹钳式两种。螺旋缠绕式是增加围角的最简单方法，因为它可以在一个滚筒上缠绕两圈或更多圈的钢丝绳，螺旋缠绕式滚筒又分为抛物线滚筒和圆锥形滚筒。抛物线滚筒进绳由直径较大处绕进滚筒上，由于倾斜力的作用，使各绳圈沿轴向运动，移动到小直径处出绳。在滚筒向某一方旋转时，只利用滚筒的一半，当绞车反转时则可应用滚筒的另一半，因此抛物线形滚筒可以正、反向工作，而圆锥形滚筒只能单向运行。螺旋缠绕式绞车的优点是结构简单，缺点是钢丝绳磨损很大，而且钢丝绳从大半径到小半径滑动的过程中，钢丝绳之间容易“咬住”，造成运转不可靠。夹钳式绞车的优点是拉力大，钢丝绳弯曲小，缺点是维护稍麻烦。夹绳弹簧质量差时易折断。1.3 设计目的 JWB型系列绞车是在老的慢速绞车基础上采用了双速多用绞车的某些结构全面改进后设计发明出来的，是一种有效矿山辅助运输设备。JWB型无极绳绞车主要用于煤矿井下工作面顺槽和采区大巷材料、设备、液压支架等辅助设备，也可用于金属矿井下巷道和地面轨道运输，适用于坡度不大有起伏变化的轨道运输。2 设计方案2.1 JWB无极绳绞车的组成JWB系列绞车主要由以下六部分组成：1).底座；2).电动机；3).制动器；4).联轴器；5) 减速箱；6）卷筒装置。外形图如图2-1JWB-75型无极绳绞车具有下列特点：1） 速度快，提高了工作效率JWB-75 型无极绳绞车工作时卷筒牵引速度4060m/min，平均速度为50m/min,是同种功用绞车产品牵引速度的3 5倍，工作效率大大提高。2） 结构简单，合理采用锥齿轮、圆柱齿轮传动、行星齿轮。只有三级传动，大大简化了机械部分传动系统，便于安装和拆卸，结构布置紧凑，合理。滚筒结构布置对称, 由一对圆柱滚子轴承支承在滚筒轴上, 支承效果好, 支承刚度高, 运行平稳。3） 实现标准化、系列化在整个机器中采用了大量的标准件，都是外购产品。另外，JWB系列绞车共有5个规格，实现了系列化设计，可以满足不同场合，不同工矿的使用。4） 操作简便，安全可靠图2-1 JWB系列绞车主要组成部分2.2绞车的工作原理由电机提供动力，通过三级传动，使行星架轴带动无极绳绞车滚筒旋转，借助钢丝绳与滚筒之间的摩擦力而达到传送重物之目的。 2.3主要构成及结构特征该绞车由底座、制动装置、变速箱、滚筒部分、联轴器等构成。 (1) 底座。由结构件焊接成整体，通过地脚螺栓与地基固定。 (2) 制动装置。制动器为电磁制动器。 (3) 变速箱。采用硬齿面齿轮。 (4) 滚筒部分。滚筒部分由主轴、卷筒及轴承等组成。 (5) 联轴器。联轴器用于联结电机和变速箱以及卷筒与行星架轴的联接。3 总体设计3.1 设计总则1.煤矿生产，安全第一。2.面向生产，力求实效，以满足客户最大实际需求。3.既考虑到运搬为主要用途，又考虑到运输、调度、回柱等一般用途。4.贯彻执行国家、部、专业的标准及有关规定。5.技术比较先进，并要求多用途。3.2 已知条件绳 速 0.78m/s牵引力 60 kN电机功率 75 kW牵引距离 1200 m设计寿命 5000 h3.3 牵引钢丝绳直径及滚筒直径的确定3.3.1钢丝绳的选择绞车的工作级别为A5钢丝绳的安全系数由式 nsg=spFtn 得nsg560000N=300000 N 式中 n钢丝绳的最小安全系数Ft 钢丝绳的额定拉力St 整条钢丝绳的破断拉力则所选型号 绳6 19 26 1550钢丝绳直径 26钢丝绳公称抗拉强度为 1550 N/mm2钢丝绳破断拉力总和 S326000 N整条钢丝绳破断拉力总和 SP= S=0.88326000 N =286880N式中 =0.88 ,拉力影响系数额定负载下的安全系数 nr=SPTmax=28688060000=4.781nr=33.3.2 滚筒参数的确定1）. 由 D0/d20 得 D020 d=520 mm取 D0=600 mm式中 D0 滚筒直径d 钢丝绳的直径2）. 确定滚筒的各个直径 (1) 滚筒缠绕直径DminDmin=D0+2d=652mm(2) 滚筒结构外径D外 D外Dmin取D外=800mm3.4 电机的选择、传动系统的确定和运动学计算3.4.1 电动机选型1）选择电机型号电动机功率 P=75kW 电动机型号 YB280S 4 同步转速 1500r/min 额定转速 1485r/min 额定功率 75kw 效 率 0.927 输出功率 P=750.927=69.525kW 重 量 650 Kg 最大转矩额定转矩=2.2 电机外形尺寸（长宽高）:1010545830 mm 电机中心高H：H = 280mm 电机轴直径长度：75140 mm2）电机输出功率的计算 已知工作机上的作用力与线速度 p=Fv1000=600.780.8083=57.90kW 3）验算电机闷车时，钢丝绳在里层的安全系数 （1）电机闷车时，钢丝绳的拉力 Tn=102Ne总vmin/60=1022.2750.808357/60 =143196.73N （2）电机闷车时，钢丝绳在里层的安全系数 nr=SpTn=286880143196.73=2.00nr=1.3 所以电机闷车时，钢丝绳是安全的。3.4.2 传动系统的确定滚筒的转速： n=60D=600.780.56=26.62 r/min 总传动比 i=148526.62=55.78采用锥齿轮-圆柱齿轮-行星齿轮三级传动示意图如图3-1图3-1 传动系统示意图传动路线:防爆电机联轴器小锥齿轮大锥齿轮小斜齿圆柱齿轮大斜齿圆柱齿轮行星齿轮行星架卷筒。3.4.3计算传动效率1各传动效率 1） 弹性联轴器效率 1=0.99 2） 圆锥滚子轴承效率 2=0.98（3对） 3） 圆柱滚子轴承 3=0.99（2对） 4） 直齿圆锥齿轮效率 4=0.96 5） 直齿圆柱齿轮传动 5=0.98 6） 行星齿轮传动 6=0.98 7） 摩擦轮钢丝绳缠绕效率 7=0.962计算传动总效率 1 =12332456 7 =0.990.9830.9920.960.980.980.9 =0.80833各级传动比分配 根据传动比分配关系知 锥-圆柱齿轮高速级传动比i134取 i1=3.80i2=2.92i3=5.034 齿轮传动设计计算4.1传动装置运动参数的计算4.1.1 各轴转速计算n1=n/i1=1485/3.8=390.8r/min n2= n1/i2=390.8/2.92=133.83r/min n3= n2/i3=133.83/5.03=26.61r/min4.1.2 各轴功率计算 P1=P 124=69.5250.990.980.96=64.76Kw P2=P135=64.760.980.98=62.19Kw P3=P236=62.190.980.98=59.73Kw4.1.3 各轴扭矩计算 T=9550p 1/n=955069.525/1485=447.11Nm T1=9550P1/n1=955064.76/390.8=1582.54Nm T2=9550P2/n2=955062.19/133.831=4437.83Nm T3=9550P3/n3=955059.73/32.43=21436.36Nm4.2锥齿轮传动设计计算4.2.1选材料、热处理方法，定精度等级。小锥齿轮材料选用：20Cr，调质 HBS=241286大锥齿轮材料选用：20Cr，表面淬火 HRC=5662 查得 Hlim=1500MPa Flim=460MPa; 采用8级精度, Ra1=Ra2=2m 4.2.2初步设计选用直齿锥齿轮 按接触强度进行初步设计，即 d1965Cm3KTu2+1 (1-0.5R)2RH2 （4.1）载荷系数 K=KAKK 使用系数KA 取 KA=1 动载系数K 推荐值 1.051.4，取 K=1.2 齿向载荷分布系数K推荐值 1.01.2 ，取K=1.1载荷系数 K=KAKK=11.21.1=1.32额定转矩 T=447.11Nm齿数比 u=3.8齿宽系数 R=0.3许用接触应力 H=0.9Hlim =0.91500=1350MPa（见表8-3-101）材料配对系数 Cm=1 （见表8-3-28） 初算结果 d196531.32447.113.82+1 (1-0.50.3)20.313502 =69.91mm4.2.3 几何尺寸计算齿数 取 z1=19 z2=uz1=3.819=72分锥角 1=tan-11u=tan-113.8=14.744 2=-1=90-14.744=75.256模数 m=d1/z1=69.91/19=3.679 m=4分度圆直径 d1=mz1=419=76mm d2=mz2=472=288mm齿宽中点分度圆直径 dm1=d1（1-0.5R） =76(1-0.50.3)=64.6mm dm2=d2（1-0.5R） =288(1-0.50.3)=244.8mm外锥距 R =d1/2sin1 =76/（2sin14.744）=149.31mm中锥距 Rm=R1-0.5R =149.31(1-0.50.3) =126.92mm齿宽 b=RR=0.3149.31=44.793mm b=46mm齿顶高 ha1=m1+x1=41=4mm ha2=m1+x2=41=4mm齿根高 hf1=m1.2-x1=41.2=4.8mm hf2=m1.2-x2=41.2=4.8mm 顶圆直径 da1=d1+2ha1cos1 =76+24cos14.744 =83.74mm da2=d2+2ha2cos2 =288+24cos75.256 = 290.04mm齿根角 f1=tan-1hf1R=tan-14.8149.31=1.84 f2=1.84齿顶角 a1=f2=1.84 a2=f1=1.84顶锥角 a1=1+a1 =14.744+1.84=16.584 a2=2+a2 =75.256+1.84=77.096根锥角 f1=1-f1 =14.744-1.84=12.904 f2=2-f2 =75.256-1.84=73.416冠顶距 Ak1=d2/2-ha1sin1 =2882-4sin14.744 =142.98m Ak2=d1/2-ha2sin2 =76/2-4sin75.256 =34.13mm安装距 取A1=147.98mmA2=104.13mm 考虑齿轮结构情况，以及轮冠距H的测量方便轮冠距 H1=A1-Ak1=5mm H2=A2-Ak2=70mm分度圆齿厚 S1=m2+2x1tan+xt1 =42=6.28mm S2=6.28mm分度圆弦齿厚 S1=S11-S126d12 =6.28(1-6.2826762) =6.273mm S2=S21-S226d22 =6.28(1-6.28262882) =6.279mm分度圆弦齿高 ha1=ha1+S12cos14d1 =4+6.282cos14.744476=4.125mm ha2=ha2+S22cos24d2 =4+6.282cos75.2564288=4.009mm当量齿数 zv1=z1/cos1 =19/cos14.744=20 zv2=z2/cos2 =72/cos75.256=283当量齿轮分度圆直径 dv1=dm1/cos1 =64.6/cos14.744 =66.7mm dv2=dm2/cos2 =244.8/cos75.256 =961.886mm齿宽中点齿顶高 ham1=ha1-12btana1 =4-1246tan1.84=3.261mm ham2=ha2-12btana2 =4-1246tan1.84=3.261mm当量齿轮顶圆直径 dva1=dv1+2ham1 =66.7+23.261=73.222mm dva2=dv2+2ham2 =961.886+23.261=969.16mm齿宽中点模数 mm=mRm/R =4126.92/149.31=3.4mm当量齿轮基圆直径 4.5 dvb1=dv1cos =66.7cos20=62.667mm dvb2=dv2cos =961.886cos20=903.877mm啮合线长度 gva =0.5dva12-dvb12+dva22-dvb22-dv1+dv22sin =0.5 73.2222-62.6772+ 969.1062-903.8772 -66.7+961.8862 sin20 =17.795mm端面重合度 va=gvammcos=17.7953.4cos20=1.774.2.4 校核接触强度 强度条件 HH 计算接触应力 H=ZHZEZZZKFmtdm1beHu2+1uKAKVKHKH （4.2）节点区域系数 ZH=2.5 弹性系数 ZE=189.8MPa 重合度系数 Z=0.88 螺旋角系数 Z=1 锥齿轮系数 ZK=0.85 名义切向力 Fmt=2000T1dm1 =2000447.1164.6=13842.41N 有效齿宽 beH=0.85b=0.8546=39.1mm 使用系数 KA=1 动载系数 KV=NK+1 其中 N=0.084z1vmt100u2u2+1 =0.084207.001003.823.82+1=0.09(vmt=d1n601000=901485601000=7.00m/s) k=(fpt-ya)cKAFmt/behCv12+Cv3 =28-2.814113842.4139.10.66+0.23 =0.277齿距极限偏差 fpt=28m 跑合量 ya=2.8 单对齿单位齿宽的刚度 c=14N/(mmm) Cv12=0.66 Cv3=0.23 则 KV=0.090.277+1=1.09 齿向载荷分布系数 KH=1.5KHbe=1.51.1=1.65 齿向载荷分配系数 KH=1.2 则 H=2.5189.80.8810.85 13843.4164.639.13.82+13.811.031.651.2 = 1206.59MPa许用接触应力 H=HlimSHminZXZLZVZR尺寸系数 ZX=1 润滑剂系数 ZL=0.975 （选100号齿轮油，运动粘度V40=100mm2/s） 速度系数 ZV=0.954 粗糙度系数 ZR=0.98 （按Ra=2m） SHmin=1 则 H=150010.9750.9540.98=1367.32MPa结论 HH 满足接触强度4.2.5齿根弯曲强度校核强度条件 FF计算齿根应力 F=KAKVKFKFFmtbeFmmnYFaYSaYYYK （4.3） 齿向载荷分布系数 KF=KH=1.65 齿向载荷分配系数 KF=1.2 有效齿宽 beF=0.85b=45.9mm 齿形系数 YFa1=2.62 YFa2=2.05 应力修正系数 YSa1=1.66 YSa2=2.06 重合度系数 Y=0.68 螺旋角系数 Y=1.0 锥齿轮系数 YK=1.0 法向模数 mmn=m1-0.5R =4(1-0.50.3)=3.4mm则 F1=11.091.651.213842.4139.13.4 2.621.660.681.01.0 =664.61MPa F2=F1YFa1YFa2YSa1YSa2 =664.612.052.062.621.66=645.32MPa许用接触应力 F=FlimYSTSFminYrelTYRrelTYX 极限应力 Flim=460MPa 安全系数 SFmin=1.4 应力修正系数 YST=2.0 齿根圆角敏感系数 YRrelT1=0.99 齿根表面状况系数 YrelT=1.04 尺寸系数 YX=1 则 F=46021.40.991.04=675.59MPa结论 F1F F2F4.2.6 齿轮各检验项目及公差值的计算 1对小齿轮: 齿距累积公差 Fp=63m 齿距极限公差 fpt=25m 齿形相对误差的公差 fc=13m 切向综合误差 Fi=Fp-1.15fc=48.05m 齿切向综合公差 fi=0.8(fpt+1.15fc)=31.96m （ 齿厚上偏差 Ess1=-22m 齿厚公差 Ts1=80m 2对大齿轮: 齿距累积公差 Fp=125m 齿距极限公差 fpt=28m 齿形相对误差的公差 fc=15m 切向综合误差 Fi=Fp-1.15fc=107.75m 齿切向综合公差 fi=0.8(fpt+1.15fc)=36.2m 齿厚上偏差 Ess2=-30m 齿厚公差 Ts2=110m 3对齿轮副接触斑点 沿齿长方向 35%65% 沿齿高方向 40%70% 最小法向侧隙 jnmin=62m 最大法向侧隙 jnmax=(Ess1+Ess2+Ts1+Ts2+Es1+Es2)cos 式中 Es1=24m, Es2=32m则 jnmax=278m 安装精度 齿圈轴向位移极限偏差 fAM=53m 轴间距极限偏差 fa=30m 轴交角极限偏差 E=30m 4.3斜齿圆柱齿轮传动设计计算4.3.1选择材料，确定和以及精度等级选择两个齿轮的材料为；大、小齿轮均为20Cr，并经调质和表面淬火，齿面硬度为5662HRC：精度等级为8级。Hlim1=Hlim2=1500MPa; FE1=FE2=900MPa, Flim1=Flim2=460MPa。4.3.2 按接触强度进行初步设计1确定中心距a（按表8-3-27公式进行设计） aCmAa(u+1)3KT1auH2 （4.4） 式中 配对材料修正系数 Cm=1螺旋角系数 Aa=476载荷系数 K=1.4 小齿轮额定转矩 T1=1582.54Nm 齿宽系数 a=0.6 齿数比 u=2.92 许用接触应力H0.9Hlim =0.91500=1350MPa 则 a476(2.92+1)31.41582.540.62.9213502165.19mm 取 a=240mm2确定模数mn mn=0.0070.02a=(0.0070.02)240 =1.684.8 取 mn=43确定齿数Z1、Z2初取螺旋角 =12 Z1=2acosmn(u+1) =2240cos124(2,92+1)=29.65取 Z1=30 Z2=uZ1=2.9230=87.6 取 Z2=894重新确定螺旋角 =cos-1mn（Z1+Z2）2a =cos-14(30+89)2240=12.732 5计算主要几何尺寸 分度圆直径 d1=mnZ1/cos =430/cos12.732=123.03mm d2=mnZ2/cos =489/cos12.732=364.98mm 齿顶圆直径 da1=d1+2ha =123.03+24=131.03mm da2=d2+2ha =364.98+24=372.98mm 端面压力角 t=tan-1tanncos =tan-1tan20cos12.732=21.06 基圆直径 db1=d1cost =123.03cos21.06=114.81mm db2=d2cost =364.98cos21.06=340.60mm 齿顶圆压力角 at1=cos-1db1da1 =cos-1114.81131.03=28.811 at2=cos-1db2da2 =cos-1340.60372.98=24.051 端面重合度 a=12Z1tanat1-tant)+Z2tanat2-tant =1230tan28.811-tan21.06+89tan24.051-tan21.06 =1.66 齿宽 b=aa=0.6240=144mm 取 b1=150mm,b2=144mm 齿宽系数 d=bd1=144123.03=1.17 纵向重合度 =bsinmn=144sin12.7324=2.53 当量齿数 Zv1=Z1/cos3 =30/cos312.732=32.33 Zv2=Z2/cos3 =89/cos312.732=91.59 基圆柱螺旋角 b=tan-1costtan =tan-1cos21.06tan12.732 =11.0974.3.3 校核齿面接触强度强度条件： HH 计算应力 H1=ZBZHZEZZFtd1bu+1uKAKVKHKH （4.5） H2=H1ZDZB （4.6）式中 名义切向力 Ft=2000T1/d1 =20001582.54/123.03 =25726.08N使用系数 KA=1 动载系数 KV=(AA+200V)-B式中 V=d1n1601000=123.03390.8601000=2.52m/s A=50+56(1.0-B) B=0.25(C-5.0)0.667 C=-0.5048lnz-1.144lnmn+2.852lnfpt+3.22（将z1=30,fpt1=28代入得C=9.42;z2=89 ，fpt2=28代入得C=8.85; 再将C值两个结果比较，并向大圆取整C=10）回代得 B=0.73,A=65.12,KV=1.26齿向载荷分布系数 KH=1.36 齿间载荷分配系数 KH=1.4 节点区域系数 ZH=2.44 重合度系数 Z=0.68 螺旋角系数 Z=0.992 弹性系数 ZE=189.8MPa 单对齿啮合系数 ZB=ZD=1 则 H1=H2=2.44189.80.680.992 25726.08123.031442.92+12.9211.261.361.4 =675.58MPa许用应力 H=HlimSHlimZNTZLZVZRZWZX （4.7）式中极限应力 Hlim=1500MPa 最小安全系数 SHlim=1. 寿命系数 ZNT=0.94（NL=60390.85000=1.17108） 润滑剂系数 ZL=1.07 （按润滑油黏度40=350mm2/s） 速度系数 ZV=0.97 粗糙度系数 ZR=0.89 齿面工作硬化系数 ZW=1 尺寸系数 ZX=1 则 H=150010.941.070.970.8911 =1302.46MPa满足 H1H，H2H4.3.4 校核齿根弯曲强度强度条件： FF计算应力: F1=FtbmnYFaYSaYYKAKVKFKF （4.8） F2=F1YFa2YSa2YFa1YSa1 （4.9）式中 齿形系数 YFa1=2.46,YFa2=2.21 应力修正系数 YSa1=1.65,YSa2=1.78 重合度系数 Y=0.25+0.75an=0.68(an=a/cos2b=1.68/cos211.097=1.73 螺旋角系数 Y=0.992 齿向载荷分布系数 KF=(KH)N=1.3 （其中N=bh21+bh+bh2,b/h应取大小齿轮中的小值,N=0.93） 齿间载荷分配系数 KF=1.4 则 F1=25726.0814442.461.650.68 0.9921.261.31.4 =280.44MPa F2 =F11.782.211.652.44=274.01MPa 许用应力 F=FlimSFlimYsTYNTYrelTYRrelTYx式中 极限应力 Flim=460MPa 安全系数 SFlim=1.25 应力修正系数 YsT=2 寿命系数 YNT=0.95 齿根圆角敏感系数 YrelT=0.99 齿根表面状况系数 YRrelT=0.99 尺寸系数 Yx=1 则 F=4601.2520.950.990.99=685.29MPa 满足 F1F F2F4.3.5 齿轮及齿轮副精度的检验项目计算1对小齿轮 （1）确定齿厚偏差代号为：8GJ GB10095-1988 （2）确定齿轮三个公差组的检验项目及公差值 第公差组检验切向综合公差Fi, Fi1=Fp1+ff1=0.090+0.022= 0.112mm 第公差组检验齿切向综合公差fi， fi1=0.6ff1+fpt1=0.60.022+ 0.028=0.030mm 第公差组检验齿向公差F=0.036mm 2对大齿轮 （1） 确定齿厚偏差代号为：8GJ GB10095-1988 （2）确定齿轮三个公差组的检验项目及公差值 第公差组检验切向综合公差Fi, Fi2=Fp2+ff2=0.068+0.028=0.096mm第公差组检验齿切向综合公差fi， fi2=0.6ff2+fpt2=0.60.022+0.028=0.030mm第公差组检验齿向公差F=0.036mm3确定齿轮副的检验项目与公差值 对齿轮，检验公法线的偏差Ew。按齿厚偏差代号KL，求得齿厚上偏差 Ess=-6fpt=-60.028=-0.168mm齿厚下偏差 Esi=-10fpt=-100.028=-0.28mm公法线平均长度上偏差 Ews=Esscos-0.72Frsin =-0.168cos20-0.720.071sin20 =-0.176mm公法线平均长度下偏差 Ewi=Esicos+0.72Frsin =-0.28cos20+0.720.075 sin20 =-245mm求得公法线长度117.183mm，跨齿数10，则公法线长度及其偏差可表示为117.183-0.088-0.175mm 对齿轮传动，检验中心距极限偏差fa，根据中心a=240mm得 fa=0.036mm；检验接触斑点，由由表8-3-64查得接触斑点沿齿高不小于30%，沿齿长不小于50%；检验齿轮副的切向综合公差FicFi= Fi1+Fi2=0.112+0.068=0.208mm检验齿切向综合公差 fic=fi1+fi2=0.030+0.030=0.060mm4.4 行星齿轮传动设计4.4.1 选取行星传动齿轮传动的传动类型和传动简图根据传动类型的工作特点选取2ZX(A) NGW 型行星齿轮传动。4.4.2 配齿计算根据传动比由表3-2直接查得各轮齿数。 太阳轮： ZA=23 行星轮： ZC=34 内齿圈： ZB=91 为提高齿轮的承载能力，采用变位齿轮，故取ZC=33 取 np=3 初选啮合角 j=ZB-ZCZA+ZC=91-3323+33=1.018 ac=23.334.4.3材料选择及热处理方式太阳轮和行星轮均采用20CrMnTi，渗碳淬火，齿面硬度5862HRC，取Hlim=1300MPa和Flim=700MPa，太阳轮和行星轮加工精度8级。 内齿圈CrMoV,调质，250280HBS4.4.4a-c 齿轮按接触强度初步计算按公式 a483E(u+1)3KTaHp2u （4.10） 1 E为齿轮副配对材料对传动尺寸的影响系数，取1 2 计算u u=ZCZA=3323=1.435 3 计算扭矩 T=T2npKc=2043.55Nm 4 K=1.22 取1.4 5 计算齿宽系数a a=2d/(u+1) ,因 d0.75 ，取d=0.5 故a=0.4116 Hp=0.9Hlim=0.91300=1170MPa 7 初定中心距 a483(1.453+1)32043.551.20.411 117021.453=179.306mm 8 计算模数 m=2aZA（u+1）=2179.306232.435=6.428mm 取 m=8mm 9 未变位时中心距a a=m2ZA+ZC=456=224mm 10 中心距变动系数Y Y=0.5ZA+ZCcoscosac-1 =0.556(cos20cos23.33-1) =0.655 11 实际中心距 a=a+Ym=224+0.6558=229.24 取 a=230mm4.4.5a c齿轮传动的主要尺寸 1实际中心距变动系数Y Y=a-am=230-2248=0.75 2实际啮合角ac ac=cos-1（aacos）=cos-1(224230cos20)=23.769 3总变位系数 X X=ZA+ZC2tan(invac-inv) =562tan20(inv23.769-inv20) =0.754 4分配变位系数 可知变位系数合适，可分配变位系数如下 Xa=0.39 Xc=0.364 5齿高变位系数Y Y=X-Y=0.754-0.75=0.004 6太阳轮a的主要尺寸 dA=mZA=823=184mm dAa=dA+2ha*+Xa-Ym =184+2(1+0.39-0.004)8 =206.176mm dAf=dA-2ha*+c*-Xam =184-2(1.25-0.39)8 =170.24mm bA=aa=0.411230=95mm 7行星轮c 的主要尺寸 dC =mZC=833=264mm dCa=dC+2ha*+Xc-Ym =264+2(1+0.364-0.004)8 =285.76mm dCf =dc-2ha*+c*-Xcm =264-2(1.25-0.364)8 =249.824mm bC =bA+5=100mm4.4.6 b-c齿轮传动的主要尺寸 a=m2ZB-ZC=491-33=232mm Y=a-am=230-2328=-0.25 bc=cos-1(aacos)=cos-1(232230cos20)=18.583 X=ZB-ZC2tan(invbc-inv) =91-332tan20(inv18.583-inv20) =-0.241 Y=X-Y=-0.241+0.25=0.009 Xb=X+Xc=-0.241+0.354=0.1234.4.7内齿圈B主要尺寸 dB =mZC=891=728mm dBa=dB-2ha*-Xb+Y-K2m K2 =0.25-0.125Xb=0.25-0.1250.123=0.23 dBa=728-21-0.123+0.009-0.238 =717.504mm bB =bA=95mm4.4.8 装配条件的验算1邻接条件 验算邻接条件，即 dAC2aacsinnp （4.11） 将各个值代入得 2642179.306sin3=310.57mm 满足邻接条件 2安装条件验算安 装条件，即 za+zbnp=23+913=38(整数) （4.12） 满足安装条件4.4.9 强度验算1a-c 传动的接触疲劳强度 （1）接触疲劳强度 HH H1=ZBZHZEZFtd1bu+1uKAKVKHKH H2=H1ZDZB 小齿轮分度圆直径 dA=184mm 分度圆上的圆周力Ft Ft=2000T2n2=20002043.55160.51=25463.21N 分度圆的圆周速度 =dAn260000=184160.5160000=1.54m/s 齿宽 b=95mm 齿数比 u=zCzA=3323=1.453 工况系数 kA kA=1 动载系数 k 计算中心轮相对于轮臂x的速度，即 x=dA(nA-nx)19100=184(160.51-32.43)19100=1.24m/s 中心轮和行星轮采用8级精度，即C=8 k=AA+200x-B 式中 B=0.25(C-5)0.667=0.2530.667=0.52 A=50+561-B=50+560.48=76.88 则 k=76.8876.88+2001.24-0.52=1.10 齿面载荷分布系数 内齿轮宽度行星齿轮分度圆直径=1001361 则 KF=KH=1 齿间载荷分配系数 KH=1.2 节点区域系数 ZH=2.5 重合度系数 Z=0.82 弹性系数 ZE=189.8MPa 则 H1= H2 =2.5189.80.8225463.21184951.453+11.4531.101.2 =701.03MPa 许用应力 H=HlimSHlimZNTZLZVZRZWZX 式中 极限应力 Hlim=1300MPa 最小安全系数 SHlim=1. 寿命系数 ZNT=0.97（NL=60160.515000=4.82107） 润滑剂系数 ZL=1.07 （按润滑油黏度40=350mm2/s） 速度系数 ZV=0.96 （按=1.54m/s） 粗糙度系数 ZR=0.89 （按Ra=2m） 齿面工作硬化系数 ZW=1 尺寸系数 ZX=1 则 H=130010.971.100.960.89=1185.14MPa 结论 H1H H2H 满足接触强度 （2）按弯曲疲劳强度校核 强度条件 FFp 计算应力: F1=FtbmYFaYSaYYKAKVKFKFFP； F2=F1YFa2YSa2YFa1YSa1 式中 齿形系数 YFa1=2.23,YFa2=2.12 应力修正系数 YSa1=1.82,YSa2=1.86 重合度系数 Y=0.25+0.75n=0.66(n=a/cos2b=1.82) 螺旋角系数 Y=1 齿向载荷分布系数 KF=1 齿间载荷分配系数 KF=1.2 F1=25463.219582.231.820.661 11.1011.21.32 =156.38MPa F2 =156.382.121.862.231.82=151.93MPa 许用应力 Fp=FlimSFlimYsTYNTYrelTYRrelTYx 式中 极限应力 Flim=700MPa 安全系数 SFlim=1.25 应力修正系数 YsT=2 寿命系数 YNT=0.97 齿根圆角敏感系数 YrelT=0.99 齿根表面状况系数 YRrelT=0.99 尺寸系数 Yx=0.98 则 Fp=7001.2520.970.990.990。98=1043.49MPa 结论 FFp 满足强度条件2b-c 传动的弯曲强度校核（1）接触疲劳强度 HH H1=ZBZHZEZFtd1bu+1uKAKVKHKH H2=H1ZDZB 小齿轮分度圆直径 dA=264mm 分度圆上的圆周力Ft Ft=2000T2n2=20002043.55160.51=25463.21N 齿宽b b=95mm 齿数比 u=zCzA=9133=2.758 工况系数 kA kA=1 动载系数 k 计算中心轮相对于轮臂x的速度，即 x=dA(nA-nx)19100=264(58.20-32.43)19100=0.36m/s 中心轮和行星轮采用8级精度，即C=8 k=AA+200x-B 式中 B=0.25(C-5)0.667=0.2530.667=0.52 A=50+561-B=50+560.48=76.88 则 k=76.8876.88+2000.36-0.52=1.06 齿面载荷分布系数 行星齿轮宽度内齿圈分度圆直径S=1.4 结论 轴的强度满足要求5.1.3 锥齿轮轴上轴承的寿命验算A B出的轴承均采用圆锥滚子轴承 32 310/T 297-1994）由于B处的轴承力比较大，校核B出即可。考虑最不利的情况，即锥齿轮的轴向力全部由B处轴承承担。 e=0.35 C=178000 FaRB=1282.222570.9N=0.0575000h结论 选择的轴承满足寿命要求5.2 大锥齿轮轴的设计计算和强度验算5.2.1 选材料预计做成齿轮轴，故选轴材料为20CrMnTi 其力学性能为 b=1100MPa,s=850MPa -1=0.27b+s=0.271100+850=526.5MPa -1=0.15b+s=0.151100+850=292.5MPa5.2.2 估算轴的直径按公式 dA3Pn 式中 d 轴径 A 与材料有关的系数， P 轴传递的功率 n 轴的转速 A=11297 d11297364.76390.859.2651.33mm 5.2.3 结构设计锥齿轮与轴采用平键联接；长度B=78mm 两端支承考虑轴向力较大，取30312轴承，内径d=60mm，宽度T=48.5mm, a=32.0mm 锥齿轮轴肩取a=4.2mm 斜齿圆柱齿轮齿根圆直径df1=98.9mm,根据轴的直径判断，制成齿轮轴。 左、右轴承轴肩轴承安装尺寸分别取轴径为67mm，90mm5.2.4 齿轮上的作用力1计算作用在大锥齿轮上的力 齿宽中点处分度圆上的圆周力为 Ftm=2000T1dm2 式中 T1 轴的转矩 dm1 齿宽中点分度圆直径 则 Ftm2=20001582.54244.8=12929.25 Fr2=Ftmtancos2 =13842.11tan20cos75.256 =1282.2N 轴向力 Fa2=Fr1=-4872.22N 2. 计算作用在斜齿圆柱齿轮上的力 圆周力 Ft=2000T1/d1 径向力 Fr1=Fttann/cos 轴向力 Fa1=Fttan 式中 d1 斜齿圆柱小齿轮分度圆 n 法面压力角，n=20 分度圆螺旋角 则 Ft =20001582.54/123.03 =25726.08N Fr1=25726.08tan20/cos12.732 =9906.40N Fa1 =25726.08tan12.732 =5812.72N 3. 计算支反力 （1）在水平面内 MAx=0 F1x=Ft1cos30-Fr1sin30=17325.59N Mx=Ftm228=12929.250.028=362.02Nm 即 RDxlAD-Ftm2lAB-F1xlAC-Mx=0 RDx=F1xlAC+Ftm2lAB+MxlAD =17325.59229+12929.2589+362020315 =17397.72N Fx=0 即 RDx+RAx-Ftm2-F1x=0 RAx=F1x+Ftm2-RDx =17325.59+12929.25-17397.72 =12857.12N （2）在垂直平面内 MAy=0 F1y=Fr1cos30+Ft1sin30 =17152.51N My=Fr228=1282.20.028=35.9Nm 即 RDylAD-F1ylAC+Fr2lAB +Fa112d1 +Fa212dm2 +My=0 RDy=F1ylAC-Fa212dm2 -Fa112d1-Fr2lAB-MylAD =17152.51229-4872.220.5244.8-1282.20.5123.03-1282.289-35900315 =9849.76N Fy=0 即 RDy+RAy+Fr2-Fr1=0 RAy=F1y-Fr2-RDy =17152.51-1282.2-9849.76 =6020.55N （3） 求合力 RA=RAx2+RAy2 =12857.122+6020.552 =14196.92N RD=RDx2+RDy2 =17397.722+9849.762 =19992.46N4.计算弯矩 （1）求水平面弯矩 MBx=RAxlAB+Mx =12857.120.089+362.02 =1506.3Nm MCx=RAxlAC-Ftm2lBC-Mx =12857.120.229 -12929.250.140-362.02 =772.18Nm （2）求垂直面弯矩 MBy=RAylAB- My =9849.760.089-35.9 =840.73Nm MCy=RDylCD=9849.760.086 =847.08Nm （3）求合成弯矩 MB=MBx2+MBy2 =1506.32+840.732 =1725.04Nm MC=MCx2+MCy2 =772.182+847.082 =1146.21Nm（4）求当量弯矩 Mac1=(MB)2+(T)2 =1725.042+0.61582.542=1969.10Nm Mac1=(MC)2+(T)2 =1146.212+0.61582.542=1488.42Nm 轴的水平弯矩图、垂直弯矩图及合成弯矩图如5.7.2 （5） 确定危险截面经过比较，根据载荷较大及截面面积较小的原则，选取截面B为危险截面 （6）校核危险截面的安全系数 1）抗弯截面模数W W=40.30cm3 2）抗扭截面系数 WP=80.6cm3 、 3）弯曲疲劳极限 -1=526.5MPa 4）扭转时的平均应力折合为应力幅的等效系数 =0.15 5）弯曲和扭转时的有效应力集中系数 k=1 kr=2.8 6）表面状态系数 =1.25 7）弯曲和扭转时的绝对尺寸影响系数 =0.66 =0.73 8）弯曲和剪切疲劳极限的综合影响系数 K=k=11.250.66=1.21 K=k=3=3.07 9）安全系数 S=-1(KMW)2+0.75(K+)TWT2 =526.5(1.212157.04/40.3)2+0.753.07+0.152217.6880.62 =5.24S=1.4结论 轴的强度满足要求5.2.5轴上键的强度验算锥齿轮上键的校核.键为 A型普通平键 型号键 221472 GB/T 1096-1979 采用双键布置 T2=1582.54Nm ,b=22mm ,h=14mm,L=72mm 根据公式 p=2000T2dkl （5.4） 式中 T2 转矩 d 轴的直径 k 键与轮毂的接触高度 平键 k=h2 C 倒角尺寸 l 键的工作长度 平键 l=L-b 则 p=20001582.54607501.5=100.48MPap 故 键的强度满足要求5.2.6轴上轴承寿命校核A D处的轴承均采用圆锥滚子轴承 30312 （GB/T 297-1994）由于D处的轴承力比较大，校核D出即可。 e=0.35 C=162000N FaRD=4872.2219992.46=0.245000h结论 选择的轴承满足寿命要求5.3 大斜齿圆柱齿轮轴的设计计算与强度校核5.3.1估算轴的直径按公式 dA3Pn 式中 d 轴径 A 与材料有关的系数，40Cr P 轴传递的功率 n 轴的转速 A=11297 d11297362.19133.8386.7575.13mm d=100mm5.3.2 轴的疲劳强度校核1计算大斜齿圆柱齿轮上的力 圆周力 Ft=2000T1/d1 径向力 Fr2=Fttann/cos 轴向力 Fa2=Fttan 式中 d1 斜齿圆柱小齿轮分度圆 n 法面压力角，n=20 分度圆螺旋角则 Ft=20004437.83/364.98 =24318.21N Fr2=24318.21tan20/cos12.732 =9074.33N Fa2=24318.21tan12.732 =5494.62N 2. 计算支反力 （1）在水平面内 MAx=0 F2x=Ft2cos30-Fr2sin30=16522.40N 即 RCxlAC-F2xlAB=0 RCx=F2xlABlAc=16522.40230329 =11550.62N Fx=0 即 RCx+RAx-F2X=0 RAx=F2x-RCx =16522.40-11550.62 =4971.78N （2）在垂直平面内 MAy=0 F2y=Fr2cos30+Ft2sin30 =20017.47N 即 RCylAC-F2ylAB+Fa212d2 =0 RCy=F2ylAB-Fa212d2 lAc =20017.47230-5494.620.5364.98329 =10946.22N Fy=0 即 RAy+RCy-F2y=0 RAy=F2y-RCy =20017.47-10946.22 =9071.25N （3） 求合力 RA=RAx2+RAy2 =4971.782+9071.252 =10344.38N RC=RCx2+RCy2 =11550.622+10946.222 =15913.41N 3 .计算弯矩 （1）求水平面弯矩 MBx=RAxlAB=4971.780.23 =1143.5m （2）求垂直面弯矩 MBy=RAylAB+Fa212d2 =9071.780.23+5494.620.50.36498 =3089.22Nm （3）合成弯矩 MB=MBx2+MBy2 =1143.512+3089.222 =3294.07Nm （4）求当量弯矩 Mac =(MB)2+(T)2 =3294.072+0.64437.832=4235.67Nm 轴的水平弯矩图、垂直弯矩图、合成弯矩图如5.7.3 （5） 确定危险截面经过比较，根据载荷较大及截面面积较小的原则，选取截面B为危险截面 （6）校核危险截面的安全系数 1）抗弯截面模数W W=86.8cm3 2）抗扭截面系数 WP=185cm3 3）弯曲疲劳极限 -1=355MPa 4）扭转时的平均应力折合为应力幅的等效系数 =0.15 5）弯曲和扭转时的有效应力集中系数 k=2.01 kr=1.88 6）表面状态系数 =1.25 7）弯曲和扭转时的绝对尺寸影响系数 =0.64 =0.72 8）弯曲和剪切疲劳极限的综合影响系数 K=k=2.011.250.64=2.51 K=k=1.881.250.72=2.09 9）安全系数 S=-1(KMW)2+0.75(K+)TWT2 =355(2.513859.71/86.8)2+0.752.09+0.155529.141852 =3.27S=1.4 结论 轴的强度满足要求5.3.3 轴上键的强度校核 1斜齿轮上键的校核.键为 A型普通平键 型号键 2816136 GB/T1096-1979 采用双键布置 T2=4437.83Nm ,b=28mm ,h=16mm,L=136mm 根据公式 p=2000T2dkl 式中 T2 转矩 d 轴的直径 k 键与轮毂的接触高度 平键 k=h2 C 倒角尺寸 l 键的工作长度 平键 l=L-b 则 p=20004437.8310081081.5=68.49MPa5000h L10hC=10660133.83（24500015913.41）103=102380.01h5000h结论 选择的轴承满足寿命要求5.4 行星轮轴的设计计算与强度校核5.4.1 初定轴的直径行星轮轴为固定心轴，不受扭矩，根据结构要求，初定轴的直径为50mm。5.4.2 轴的疲劳强度校核1.求支反力 太阳轮对行星轮的作用力 Ftac=2000TanpdaKc （5.6） 由于啮合角相等，内齿圈对行星轮的作用力 Ftbc=Ftac 式中 Ta=T2ZAZA+ZB T2 扭矩 np 行星轮数目 da 太阳轮分度圆直径 Kc 载荷分配不均匀系数 则 Ta=4437.8323114=1115.52Nm Ftac=Ftac=9700.18N RAx=RBx=0.5Ftac+Ftbc=9700.18N 2.求弯矩 MC=RAxlAC=9700.180.0345=334.66Nm 3.求当量弯矩 Mac =(MC)2+(T)2 =334.662+0.61115.522=748.32Nm 受力图和弯矩图5.7.4 4.确定危险截面 截面C为危险截面 5.校核危险截面安全系数根据公式 S=2-1K+M/WS （5.7） 式中 -1材料的弯曲疲劳极限,按40Cr K 弯曲疲劳极限的综合影响系数 弯曲的平均应力折合为应力幅的等效系数 M 弯矩 W 抗弯截面模数 则 S=2355(3.66+0.25）334.66/12.5=6.78S 满足强度要求5.4.3 轴承的寿命校核选用的是调心滚子轴承22110C (GB/T 288-1994) C=83700N RA=9700.18N L10h=10660n(CP)h 式中 L10h 以小时数（h）表示的轴承的基本额定寿命 n 轴承工作转速，r/min C 基本额定动载荷，N （表） P 当量动载荷，N 寿命系数则 L10h=10660236.74（837009700.18）103=92767.53h5000h结论 选择的轴承满足寿命要求5.5 行星架轴的设计计算与强度校核5.5.1估算轴的直径按公式 dA3Pn 式中 d 轴径 A 与材料有关的系数，40Cr P 轴传递的功率 n 轴的转速 A=11297 d11297362.1926.61147.6128.73mm d=140mm5.5.2 渐开线花键的强度校核1花键尺寸计算 分度圆直径D D=mZ=522=110mm 基圆直径 Db=mZcosD=110cos30=95.26mm齿距 p=m=5=15.71mm 内花键大径基本尺寸 Dei=mZ+1.5=522+1.5=117.5mm 内花键大径下偏差 0 内花键大径公差 0.32mm内花键渐开线终止圆直径最小值 DFimin=mZ+1.5+2CF =523.5+1=118.5m内花键小径基本尺寸 Dii=DFemax+2CF =104.52+1=105.52mm 内花键小径极限偏差+350 0 基本齿槽宽 E=0.5m=0.55=7.85mm作用齿槽宽最小值 EVmin=0.5m=0.55=7.85mm 实际齿槽宽最大值 Emax=EVmin+T+ =7.85+0.056=7.905mm 实际齿槽宽最小值 Emin=EVmin+=7.85+0.036=7.886mm 作用齿槽宽最大值 EVmax=Emax-=7.905-0.036=7.869mm 外花键作用齿厚上偏差 esV=0 (见表3-3-29) 外花键大径基本尺寸 Dee=mZ+1=523=115mm外花键大径上偏差 esVtanD=0 外花键大径公差+350 0 外花键渐开线起始圆直径最大值 DFemax =0.5Db2.+(0.5DsinD-hs-0.5esVtanDsinD)2 =104.52mm 外花键小径基本尺寸 Die=mZ-1.5=520.5=102.5mm外花键小径上偏差 esVtanD=0 外花键小径公差 0.32mm 基本齿厚 S=0.5m=0.55=7.85mm作用齿厚最大值 SVmax=S+esV=7.85 实际齿厚最小值 Smin=SVmax-T+ =7.85-0.056=7.794mm 实际齿厚最大值 Smax=SVmax-=7.85-0.036=7.814mm 作用齿厚最小值 SVmin= Smin+=7.794+0.036=7.83mm 齿形裕度 CF=0.1m=0.5mm2花键的挤压强度校核 p=2T3ZhlDmp 式中 T3 转矩 各齿载荷不均匀系数 一般取0.70.8 Z 齿数 Dm 平均直径 h 齿的工作高度 l 齿的工作长度 m 模数 则 p=200021436.360.75225104110=45.43MPap 花键满足强度要求5.6 卷筒轴的设计5.6.1估算轴的直径按公式 dA3Pn 式中 d 轴径 A 与材料有关的系数，40Cr P 轴传递的功率 n 轴的转速 A=11297 d11297362.190.990.9826.61147.14127.43mm 取 d=150mm5.6.2 轴的疲劳强度校核1 .求支反力 钢丝绳牵引力 F=60000N RA=RB=30000N 2. 求弯矩 MA=MB=RA0.5 lAB =300000.5874 =13110Nm3求当量弯矩 Mac =(MA)2+(T)2 =131102+0.621436.362=18356.69Nm 受力及弯矩图如5.7.5 4. 确定危险截面 由弯矩图得，c截面为危险截面 5校核危险截面安全系数1）抗弯截面模数W W=329.44cm3 2）抗扭截面系数 WP=660.78cm3 3）弯曲疲劳极限 -1=355MPa 4）扭转时的平均应力折合为应力幅的等效系数 =0.15 5）弯曲和扭转时的有效应力集中系数 k=2.01 kr=1.88 6）表面状态系数 =1.25 7）弯曲和扭转时的绝对尺寸影响系数 =0.60 =0.68 8）弯曲和剪切疲劳极限的综合影响系数 K=k=2.011.250.60=2.68 K=k=1.881.250.68=2.21 9）安全系数 S=-1(KMW)2+0.75(K+)TWT2 =2.03S=1.4 结论 轴的强度满足要求5.6.3 轴承的寿命校核选用的是圆柱滚子轴承N2224 (GB/T 283-1994) C=332000N RA=30000N L10h=10660n(CP)h 式中 L10h 以小时数（h）表示的轴承的基本额定寿命 n 轴承工作转速，r/min C 基本额定动载荷，N （表） P 当量动载荷，N 寿命系数则 L10h=1066026.61（33200030000）103=1876635.35h5000h结论 选择的轴承满足寿命要求5.6.4 键的校核.键为 A型普通平键 型号键4022180 GB/T1096-1979 采用双键布置 T2=21436.36Nm ,b=40mm ,h=22mm,L=180mm 根据公式 p=2000T2dkl 式中 T2 转矩 d 轴的直径 k 键与轮毂的接触高度 平键 k=h2 C 倒角尺寸 l 键的工作长度 平键 l=L-b 则 p=200021436.36160111401.5=112.72MPap 键的强度满足5.7 轴的受力简图 5.7.1小锥齿轮受力简图5.7.2大锥齿轮受力简图 5.7.3大斜齿轮受力简图5.7.4行星轮轴受力简图5.7.5卷筒轴受力简图6 JWB-75型无极绳绞车使用说明书61使用范围本绞车用于煤矿井下工作面顺槽和采区大巷材料、设备、液压支架等辅助运输，也可用于金属矿井下巷道和地面轨道运输，适用于坡度不大有起伏变化的轨道运输，最大起伏度16，适用运输距离约为2000m。62主要技术规范请详参见附表一主要技术参数表。63绞车的润滑与密封机器的润滑不仅关系着机器的正常工作，而且直接影响着机器的寿命，因此必须及时地更换和补充润滑油。润滑油的优质必须符合要求，不得混入灰尘、污物、铁屑及水等杂质。闭式齿轮传动润滑油采用工业齿轮油250号（SY1172-77S）。减速箱内最高油面不超过大锥齿轮直径的三分之一。闭式传动箱内的滚动轴承均为溅油润滑。开式齿轮传动及滚动轴承均采用钙钠基脂润滑脂（SY140377），各滚动轴承内加入润滑脂加入量不得超过容量的三分之二，每隔36个月加油或更换一次。对于新的或大修后的绞车，在运转半个月后必须更换变速箱内的润滑油并进行清洗，以除去传动零件磨落的金属细屑。变速箱剖分面及各密封面，密封后均不许漏油，在各密封面涂密封胶，或水玻璃。64绞车的装配、调整、及试运转绞车在装配前应清查零件数量，并将所有零件修净毛刺，清洗干净，严防铁屑、灰尘带入绞车内部，所有滚动轴承必须在油中加热后（一般加热温度为120150C）进行装配。不得硬打硬砸，以防零件受力变形。绞车应先进行部装，然后进行总装。总装前应先装成变速箱部分，卷筒装置部分两个主要部分。6.4.1变速箱在装配该部分时，可按以下程序进行装配。(1)先将齿轮轴、花键轴等各组件分别进行装配，然后依次将各组件装在箱体上。(2)锥齿轮轴正确的轴向位置调整可借助套杯与箱体间的调整垫片来实现，大锥齿轮及二轴上的斜齿圆柱齿轮正确的轴向位置调整可借助于二轴组件的闷盖，轴套和轴承支架以及箱体间的两组调整垫片来实现。花键轴组件上零件的正确轴向位置由轴端闷盖与箱体间的调整垫片来实现。(3)锥齿轮传动保证侧隙为0.16，闭式斜齿圆柱齿轮传动保证侧隙为0.230。(4)圆锥滚子轴承的轴向游隙为0.0500.100，圆柱滚子轴承的轴向游隙为0.0500.080。(5)用图色法来检验斑点，锥齿轮按齿高和齿宽接触斑点都不小于60%，闭式斜齿圆柱齿轮接触斑点沿齿高和齿宽均不小于45%，开式圆柱齿轮接触斑点沿吃齿高不小于40%，沿齿宽不小于50%。(6)最后将通气塞、透视盖、油标及其它零部件装好，并按第四节规定加注润滑油和润滑脂。6.4.2 卷筒装置该部分主要包括卷筒、圆柱滚子轴承、卷筒轴、防尘圈、轴承盖、轴承支架、圆盘等。轴承的轴向游隙为0.0500.080。6.4.3总装各部分位置参见图1，其总装程序如下：先把变速箱部件和制动闸同时放在底座上，然后把卷筒装置装在底座上，其后再装电动机，加减电动机脚下的调整垫片调整电动机的中心高，电动机轴心线与齿轮轴中心线的同轴度允差为0.32mm，两中心线的倾角不得大于40分。制动闸内径与联轴器外径之间间隙调整到1.52.5mm，也可根据需要进行调整，但应使施闸时灵活，无卡住现象。所有部件装配、调整完毕后，将各部分连接螺栓紧固。6.4.4 空负荷试运转新的或大修后的绞车，空负荷试运转在转配完毕后进行。试运转应遵守操作规程，试运转前必须先确认变速箱润滑油清洁度符合要求。先用手动或电机点动，机器正常运转后方可进行空负荷试运转试验。空负荷试运转时，正反转30分钟并达到下列要求。(1) 运转正常，无冲击性噪声；(2) 各部分润滑良好，无渗漏现象；(3) 高温度不超过80C，最高温升不超过40C；(4) 各部无松动现象；(5) 制动闸应制动灵活、可靠。空负荷试车后，焊轴支承挡块，电动机挡块。6.4.5 负荷试运转空负荷试运转后方可进行负荷试运转，负荷试运转应注意以下事项：(1)绞车的出绳方向，钢丝绳上出绳既钢丝绳引出部分应在卷筒上面，不宜在卷筒下面出绳；(2)负荷试车如在现场,可在井下利用工作面支柱作为负荷,负荷试车时必须将绞车牢固地固定在底板上;(3)试车负荷应逐次增加,并且应当注意钢丝绳的最大牵引力为80kN;(4)负荷试车时,变速箱内油温最高不超过90C,最高温升不超过80C.(5)制动闸制动灵活、可靠;(6)每次负荷试车后检查钢丝绳及机器各部不得有残余变形，并应符合试运转中第13条规定；(7)负荷试车完毕后必须更换变速箱中的润滑油。6.5 绞车的固定方法和安设位置6.5.1 绞车在现场的固定方法本绞车固定方法见图6.1。绞车固定后应牢固、可靠，不得有松动现象。6.6 绞车的操作规程绞车的操作比较简单，按动按钮、启闭电动机、操纵电磁制动闸的控制按钮及可实现绞车的整个动作过程。6.6.1 工作前的注意事项(1)检查钢丝绳，钢丝绳应符合煤矿安全生产试行规程的有关规定；(2)检查钢丝绳固结的是否可靠，各连结螺栓是否紧固，绞车安装是否牢固等；(3)检查变速箱内润滑油是否充足，发现不足应适量补充，但润滑油油质不得随意更改；(4)检查制动闸；(5)检查使用电源、电缆和电器设备的接线是否正确，接地是否安全，是否有漏电现象。6.6.2 工作时应遵守下列规定:(1)绞车司机应集中精力，注意倾听信号；(2)绞车工作时，注意整理钢丝绳，使其缠绕整齐，工作钢丝绳不能全部放完，在卷筒上至少保留三圈；(3)绞车工作时可能发生的故障及消除方法如下表：序号故障现象故障可能发生原因故障消除方法1开机时电机不转或发出叫声载荷过大或接触不良停止运转使电机反转卸载或检查接线2机器跳动安装不牢或地基不平整理地板或从新安装3机器声音不正常零件装配不正确,零件磨损过多或连接松动停车检查6.6.3 工作后应注意的事项(1)工作结束后应将钢丝绳整齐的缠绕在卷筒上，切断电源，关闭开关；(2)除机器上的灰尘、杂物；(3)交接班时必须把本班发现的不正常情况向下一班交代清楚，以便及时消除不正常情况。6.7绞车的维护、拆卸与修理(1)绞车的操纵人员必须严格遵守操作规程。(2)绞车必须按第四节的规定及时加注润滑油。(3)绞车如长时间搁置不用，必须选择干燥的地方存放，防止电器受潮，绞车的裸露部分应涂以保护油，各摩擦部分涂上润滑油。(4)有关电机维护可详见随机文件YB280S-4W型防爆异步电动机产品说明书。(5)绞车的拆卸次序和装配相反，绞车应先拆成部件，然后再将各部件进行拆卸，拆卸时应先将卷筒装置、电机与底座的连接螺栓松开，并将该两部分拆掉，然后再将变速箱从底座上拆掉。变速箱的拆卸是先将上箱体上的调速装置部分零件、油标通气器等拆下，将将上箱体拆下，然后把齿轮轴、二轴、花键轴各组件拆下，最后将所有零件拆除。(6)拆卸绞车各部时，应注意各部位的调整垫片的数量和厚度，以便在重新装配各部件时，保证绞车原有的装配精度，特别要注意锥齿轮副的调整垫片不得任意增减。(7)绞车在拆卸过程中，严禁用锤硬打硬砸，必须小心进行，不得损坏零件或碰伤零件表面。(8)绞车应按实际情况，有计划安排小修、中修、大修计划，绞车的修理周期、修理内容、修理场所根据煤炭部制定的煤炭工业设计规范一文中的有关章节，作如下规定：1）小修：小修周期为三个月，一般在现场进行，主要调整更换钢丝绳和紧固连接件，并消除故障，补充或更换润滑油，清洗绞车外表灰尘等；2）中修：中修周期一般为9个月，中修一般在矿机厂进行，主要任务是全部拆开绞车各部分，清洗后检查磨损程度，更换已磨损的零件，消除小修时不能消除的故障，更换机器各部润滑油，恢复绞车工作能力和正常状况，中修后应进行试运转；3）大修：大修周期为18个月，大修一般在矿机厂进行，其主要内容是拆开绞车全部零件清洗和检查一切零件，修复或用新的零件来替换已磨损的零件，全部恢复绞车的工作能力和正常状况。大修后应进行试运转，并进行油漆更新。图6-1绞车地角螺栓固定示意图附表一 主要技术参数表附表二 齿轮一览表 序号名称材料齿数模数压力角齿宽精 度等 级1齿 轮轴20Cr194204682大 锥齿 轮20Cr724204683齿 轮轴20CrMnTi30421.0601508 4大 斜齿 轮20Cr89421.06014485太 阳轮20CrMnTi238209586行 星轮20CrMnTi3382010087 内 齿圈CrMoV91 8209587 结论这次毕业设计，通过查阅相关资料，我了解了绞车特别是无极绳绞车的发展历史及其现状，并对绞车的工作原理有了一定的认识。在设计JWB-75型无极绳绞车的过程中，我分析了该绞车的传动结构，掌握了传动系统及总体布局设计。JWB-75型无极绳绞车在传动系统方面实现了一个封闭传动路线，更加有利于绞车的可靠制动和方便操作。另外初步解决了该绞车中行星轮浮动的问题。在这次毕业设计中，我经常去图书馆借阅有关的图书资料，还上网查阅有关绞车方面的最新资料和信息，综合了大学四年所学的机械设计方面的知识完成了这次设计任务。通过这一过程，我体会到了如何才能成为一名合格的机械设计工作者；在毕业实习过程中，我也更加理解了理论联系实际的重要性。这次绞车的设计主体部分是减速器的设计，另外还有卷筒、底座的设计。JWB-75型无极绳绞车的减速器是一个三级减速器。由一对直齿锥齿轮传动，一级斜齿圆柱传动和一个行星轮传动组成。设计中有涉及到齿轮、轴的大量计算以及轴承、键的设计选择和校核，另外还有电机、制动器和各种螺纹连接件的选择。在反复的计算和校核过程中，选择最佳方案。本次设计，我充分利用各方面的资料，借鉴前人的经验，。在指导老师和同学的帮助下，顺利的完成了毕业设计。这次设计使我学到了许多机械设计方面的知识，受益匪浅。由于本人能力有限，还望各位老师多多批评和指正。参考文献1 卜炎. 机械传动装置设计手册. 北京：机械工业出版社，19992 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，20013 机械工程手册编委会. 机械工程手册（第二版）.机械工业出版社，19964 吴宗泽. 机械设计实用手册（第二版）.北京：化学工业出版社，20035 吴宗泽. 机械设计师手册.北京：机械工业出版社，20026 蔡春源. 新编机械设计手册. 沈阳：辽宁科学技术出版社，19937 机械工程手册编委会. 机械工程手册（第二版）.机械工业出版社，19968 刘鸿文. 材料力学. 武汉：华中科技大学出版社，20049 郝桐生. 理论力学. 北京：高等教育出版社，200510 中国矿业大学机械制图教材编写组.画法几何及机械制图. 徐州：中国矿业大学出版社，200211 甘永立. 几何量公差与检测.上海：上海科学技术出版社，200112 程志红. 唐大放. 机械设计课程上机与设计. 徐州：东南大学出版社，200613 程志红.机械设计. 徐州：东南大学出版社，200614 王绍定. 矿用小绞车. 北京：煤炭工业出版社，198115 洪晓华. 矿井运输提升. 徐州：中国矿业大学出版社，200516 张树森. 机械制造工程学. 沈阳：东北大学出版社，200117 庄宗元. AutoCAD 2004 使用教程.徐州：中国矿业大学出版社，200418 韩正铜. 机械精度设计与检测.徐州：中国矿业大学出版社，200719 闫吉领. 无极绳绞车在矿井生产中的应用及改进措施. 煤矿机械，2002,51-5220 贾双春，徐剑民，王建慧. 无极绳绞车运输系统的改进和优化. 实用技术，2007,21-2221 李炳文，刘海平. 新型无极绳绞车制动系统的研究. 矿上机械，200622 刘海平，李炳文，曲利等. 新型矿用无极绳绞车传动系统的设计. 煤矿机械，2005，161-16223 马明祥，胡刚，王少华. 无极绳绞车运输的工艺改进. 煤矿现代化，2007，47 24 Griffin, B.A High performance winch and synthetic rope systems for workboats, tug boats ,and commercial marine applications. Oceans 04 MTS/IEEE Techno-Ocean, p 1900-3 Vol 4,200425 Wang liwen. Analyzing of the winch capacity by using adjustable hydrodynamic elements.Procdddings of the srd international Symposium on Fluid Power Transmission and Control(ISFP99), p 512-515,1999翻译部分英文原文Reducing Commercial Fishing Deck Hazards with Engineering Solutions for Winch DesignJennifer M. Lincoln a,* , Devin L. Lucas a , Robert W. McKibbin b,Chelsea C. Woodward b, John E. Bevan ba NIOSH Alaska Field Station, Anchorage Alaskab NIOSH Spokane Research Lab, Spokane WashingtonAvailable online 20 March 2008AbstractIntroduction: The majority (67%) of hospitalized injuries among Alaska commercial fishermen are associated with deck machinery. This paper describes the “Prevention Through Design” process to mitigate one serious machinery entanglement hazard posed by a capstan deck winch. Methods: After observing that the capstan winch provides no entanglement protection and the hydraulic controls are usually out of reach of the entangled person, NIOSH personnel met with fishermen and winch manufacturers to discuss various design solutions to mitigate these hazards. Results: An emergency-stop (“e-stop”) system was developed that incorporated a momentary contact button that when pushed, switches a safety-relay that deenergizes the solenoid of an electro-hydraulic valve stopping the rotating winch. The vessel owners that had the e-stop installed enthusiastically recommend it to other fishermen. NIOSH entered into a Proprietary Technology Licensing Agreement with a company to develop the system for commercial use. Conclusions: This is an example of a practical engineering control that effectively protects workers from a hazardous piece of equipment by preventing injuries due to entanglement. This solution could reduce these types of debilitating injuries and fatalities in this industry. 2008 National Safety Council and Elsevier Ltd. All rights reserved.Keywords: Prevention through design; PtD; Commercial fishing; Entanglement hazards; Engineering solution; Injury prevention1. IntroductionCommercial fishing is the most dangerous occupation in the United States. In 2006, 51 commercial fishermen were killed on the job resulting in an occupational fatality rate of 142 per 100,000 workers, the highest rate for any occupation in the United States and 36 times higher than the average fatality rate for all U.S. workers (Bureau of Labor Statistics BLS, 2007). During 19942004, 641 commercial fishermen died in the United States. Of these, 332 (52%) died after their vessel sank, and another 184 (29%) fatalities were due to falls overboard. The remainder of the fatalities were due to a variety of causes, including deck injuries (51, 8%; Dickey & Ellis, 2006). These fatal deck injuries are from machinery and fishing gear. In Alaska, fatal deck injuries are even more prevalent, accounting for 12% of all fatalities during 20002006 (Commercial Fishing Incident Database CFID, 2007). Prevention efforts should emphasize preventing the loss of life due to the loss of a vessel, falls overboard, and injuries on deck. This paper focuses on the efforts to prevent injuries on deck, including the redesign of machinery or the retrofitting of safety features on fishing machinery and equipment.An important issue to address is jurisdiction of regulatory agencies on uninspected commercial fishing vessels. Both the Occupational Safety and Health Administration (OSHA) and the U.S. Coast Guard (USCG) have authority over the safety of employees onboard commercial fishing vessels. The USCG is the lead agency; OSHA only has jurisdiction out to 3 miles from shore. In addition OSHA is precluded from enforcement with respect to working conditions regulated by other federal agencies. Therefore, the promulgation of safety regulations by the USCG preempts OSHA with respect to those working conditions specifically addressed by Coast Guard regulations. The USCG does have regulations in regard to machine guarding, but the extent to which these are enforced is limited.In Alaska, most fatalities in the commercial fishing industry are also due to the loss of a fishing vessel. However, injury epidemiologists at the National Institute for Occupational Safety and Health (NIOSH) Alaska Field Station have shown that most (67%) severe nonfatal injuries occur on deck during the deployment and retrieval of fishing gear (Thomas, Lincoln, Husberg, & Conway, 2001). Severe nonfatal injuries are defined as those requiring hospitalization and include lacerations, broken bones, severe head injuries, and smashed limbs. The deck of a fishing boat is a slippery, constantly moving work platform that is often congested with machinery and fishing equipment. Many of the deck machines used on commercial fishing vessels lack adequate guarding and safety features and entanglement is a particular hazard. NIOSH found that during 19912002, 798 fishermen were hospitalized for severe nonfatal injuries, which is equivalent to an annual rate of 410 per 100,000 full-time equivalent fishermen. Of these injuries, 23% were due to being entangled or struck by lines or gear, or being trapped in a winch, pulley, or other deck equipment (Lincoln, Husberg, & Mode, 2006). Experts have recommended that vessel machinery be redesigned or retrofitted with safety features to prevent these types of injuries (Husberg, Lincoln, & Conway, 2001; Burgess, 2001). The purpose of this paper is to describe the “Prevention Through Design” activities we completed in order to mitigate one machinery entanglement hazard posed by a deck winch commonly found on commercial fishing vessels.2. MethodsNIOSH is the federal agency responsible for conducting research and making recommendations to improve the safety and health of workers in the United States. The NIOSH Alaska Field Station has worked in the area of commercial fishing safety since 1991. After working with our industry partners on several safety issues, NIOSH developed the Deck Safety Intervention Project, which began in October 2000. Goals of this project were to determine when and where deck injuries occur and to develop intervention strategies. These intervention strategies included engineering designs with industry input to eliminate or lessen the risks deck machines pose. In 2002, NIOSH injury epidemiologists met with fishermen across Alaska to discuss the problem that 67% of severe hospitalized injuries were a result of deck machinery, with the hope that practical solutions could be developed to preventthese injuries from occurring. One significant entanglement problem fishermen identified was that posed by the capstan-type winches typically found on purse-seine vessels. Vessels fishing with purse-seine gear are generally about 50- feet long and are accompanied by a skiff, which is integral to the fishing operation. The seine is a large net with thick mesh and small openings. It is not designed to catch fish in the mesh; rather, the net acts as a large cage trapping the fish. The seine is set by using the skiff to pull one end of the net off the stern of the vessel. The skiff is brought around so that the seine makes a half circle with the skiff on one end and the vessel on the other end. Both ends are towed equally to gather a school of fish. Then, the skiff pulls its end over to the vessel completing the circle. At this point, a line, which is threaded through rings at the bottom of the net, is wrapped around the drum of the capstan deck winch and pulled in. This line, the “purse-line,” draws the bottom of the net together making a large “bag” or purse. The fish trapped in the purse are brought onboard and dumped in the fish hold.The deck winch on purse-seine vessels is a powerful (up to 1000 foot-pounds of torque) capstan winch, usually mounted in the center of the deck near the wheelhouse. Its drum rotates while the crew is working on deck. Fishermen who lose their balance or are inattentive can become entangled in the purseseine as the line is winding around the drum. Crushing injuries to the hand or arm, and in some cases fatalities if the head or torso is caught, are the results. The winch hydraulic motor controls are usually located on the bulkhead just forward of the winch. Unfortunately, they are usually out of reach of the entangled fisherman to stop the drum rotation. In June 2002, tragedy struck the captain of a commercial purseseine vessel fishing for salmon in Prince William Sound, Alaska. The crew had just set the seine around a school of fish. After drifting with the net open for about 15 minutes, the captain called for the skiff-man to close the net, which was done without incident. Then, as the captain wrapped the purse line around the drum of the deck winch to close the bottom of the seine, he reached across the winch, brushing his raincoat sleeve against the moving line. The sleeve caught between the line and the rotating drum. He called for help as he was being wrapped into the spinning drum. Two crewmen charged across the deck from the stern where they tried to work the hydraulic winch controls, located on the backside of the pilot-house. However, the captain had gone around the winch three times before it was stopped. Despite the best efforts of the crew and others nearby, the captain died the next day at an Anchorage hospital from multiple traumatic injuries. One year later in October 2003, another purse-seine captain was killed in the same type of capstan deck winch near Homer, Alaska. Working alone during an opening for rockfish, he was using the deck winch to lower a bag of fish that he had just weighed into the fish hold. Evidence suggested that the captains right hand got too close to the rotating drum and his coat sleeve got caught in the line. Entangled in the deck winch, the controls were out of reach and the rotating winch could not be stopped. The captains body was found mangled and wrapped repeatedly around the capstan several days later when his wife reported him overdue to the U.S. Coast Guard. NIOSH investigations identified two major safety hazards: no entanglement protection is provided by the winch and the hydraulic controls are usually out of reach of the person who is entangled. Injury epidemiologists from the NIOSH Alaska Field Station partnered with engineers at the NIOSH Spokane Research Laboratory to design a practical engineering solution to mitigate these safety hazards. This NIOSH team of epidemiologists and engineers met with vessel owners, purse-seine fishermen, and winch manufacturers to discuss the various safety design options. It was quickly realized that standard machine guarding and “dead-man switch” solutions were either not feasible or applicable to the typical machine use. Pressure mats for the dead-man switch would be subject to false signals from lines being coiled on deck, or from fish as the net is dumped. Physical guards are impractical because lines are fed onto the winch from virtually any angle using the fixed winch horns in combination with the rotating drums (complicated by the fact that both drums are sometimes used at the same time with lines from two directions). Rain or ocean spray would interfere with light curtains. Design considerations also favored systems that would be simple, affordable, unobtrusive, applicable to various winch models, use off-the-shelf components, not disable other vessel functions (such as the rudder or anchor winch), not interfere with normal fishing operations if the emergency-stop system failed, and most importantly have the capability to be activated by the person being entangled in time to prevent serious injury (McKibbin & Woodward, 2006).3. ResultsThe engineering design solution that was developed was an emergency-stop (“e-stop”) system that incorporated a robust, low-profile, momentary contact button mounted on the top portside winch horn. This location was the preferredmounting spot for a fisherman pursing the net from the starboard side of the vessel (the most common scenario). When pushed, the button energizes a safety-relay that in turn deenergizes the solenoid of an electro-hydraulic valve. This valve, plumbed between the manual valve that controls winch rotation and the winch drive motor, closes the flow of hydraulic oil to and from the winch motor and locks the capstan drums in place. If the emergency switch is pressed in a timely manner and the hydraulic motor does not have significant wear in the vane seals, the winch drums will lock almost instantaneously. When the electro-hydraulic valve is de-energized, the valve-spool shifts to the default position to block the oil flow to and from the winch motor. This functions as a hydraulic brake on the winch enabling a Category 0 stop. If the electric power to the electrohydraulic valve is turned off for any reason, the hydraulic fluid flow stops and the winch will not operate until the power is returned and the e-stop circuit is reset. If the winch drum has significant external rotational force acting upon it, and the motor seals are in need of repair, back-spooling of the winch drums could occur after the hydraulic valve closes both pump and tank motor ports. The back-spooling drums could effectively unwind the victim from the entanglement. The time it takes the victim to strike the emergency switch is thus the controlling factor in arresting the drum rotation, not the valve response time. With the drums turning at a typical working speed (4060 rpm), and considering a typical humanresponse time (less than 0.5 sec), the winch drums could be stopped in less than 180 of rotation sufficient to limit serious entanglement injury. A control box containing the safety-relay, pilot lights to indicate system status, and a system reset button, is mounted adjacent to the winch directional control valve. In the event someone becomes entangled, it is important that the winch directional control valve be returned to neutral before the e-stop system is reset; otherwise, further entanglement or injury could result by the rotating drum. On purse seine vessels, the manual winch directional-control valve is almost always mounted on the rear bulkhead of the vessel, usually about 5- to 6-feet forward of the capstan winch. The normal function of the manual valve precludes having a self-centering spool. The fisherman typically shifts the spool for either a forward or reverse winch operation, then leaves the winch running to execute his/her fishing duties. If there is a winch entanglement and the e-stop is used to stop the drum rotation, the safety relay requires a manual reset. If the prototype system was reset without verifying that the manual valves were in the neutral position, the winch drums may start to rotate again unexpectedly. The position of the reset function was moved from the position of the e-stop switch to the proximity of the manual valves. In the described scenario, this operating location removes the reset-switch operator from the possible hazardous area near the rotating winch drums. Yet, he/she is close enough to make a visual inspection of the clearance around the winch and to be able to observe the position of the manual valves, before resetting the system. This design lends itself to the development of more advanced systems. Additional buttons, easily wired in parallel to the winch e-stop button, can be placed in other locations on the vessel such as in the wheelhouse or along the gunwales, leaving the reset switch at the manual winch controls. The e-stop system may be applied to other types of deck machinery, and may eventually incorporate wireless or voice activation features. A fishing vessel owner/operator home-ported in Seattle, Washington worked with NIOSH on the design and installation of this e-stop system, which was successfully tested during the 20052007 Southeast Alaska salmon seasons. This vessels crew praised the device as a significant safety and, surprisingly, productivity improvement. As of this writing, they continue to use the winch e-stop system. The crewmembers have reported that although they have not had to use it in an emergency, it has come in handy when the purse-line gets knotted up while operating. This capstan winch e-stop system has been installed by NIOSH on two other seiners that operate in Alaska. Feedback from a qualitative evaluation of the e-stop system was completed. We asked the skippers who are currently using the system questions in regard to its acceptability, use in an emergency, its reliability, durability, possible design improvements, and examples of how it is making their fishing more efficient. The other fishing crews also like having the e-stop on their vessels. They believe the e-stop is very helpful, very important, and would enthusiastically recommend it to other fishermen. None of them have had to use it in an emergency. They did indicate that they found it increases efficiency because they are able to stop the winch faster when something is wrong with the lines. On rare occasions, the button is sometimes pushed unintentionally by someones hand or elbow, but the system can be quickly reset so as not to be a problem. Note that the NIOSH designed e-stop does not reset automatically, and it cannot be reset by pushing the e-stop button a second time (i.e. toggling). It requires that someone manually push the reset button mounted in the control box on the rear bulkhead. Fishermen who have used the system are so pleased with it that they have cited other pieces of machinery on their vessels that could be made safer with similar e-stop buttons. The e-stop has received strong support at industry trade shows. It has been demonstrated at Pacific Marine Expo in Seattle in 2005, 2006 and 2007, the largest commercial fishing trade show in the United States, and at Comfish Alaska in Kodiak in 2007. The e-stop has received an overwhelming positive response from fishermen, with many fishing vessel owners and operators requesting information on how to obtain the device for their vessels. NIOSH staff have also produced and distributed copies of a deck safety awareness DVD that illustrates the entanglement hazard and e-stop solution to the hazard (NIOSH, 2007). Over 100 DVDs have been distributed to fishermen, marine safety organizations, and government agencies in the nine months since production. In the next year, a control technology publication will be written to increase the distribution and impact of this e-stop solution. The NIOSH intervention design has been licensed to a manufacturer. In November 2007, NIOSH entered into a Proprietary Technology Licensing Agreement with Emerald Marine Products, LLC (Seattle, WA) to develop the e-stop system for commercial use. Emerald Marine Products, in cooperation with Kolstrand Marine Supply (Seattle, WA), will be manufacturing a similar system based on the NIOSH research. They will be selling the e-stop exclusively through their distributor Go2Marine.4. ConclusionsThe e-stop is an example of a practical engineering control or physical modification that can be implemented to protect workers from a hazardous piece of equipment such as the capstan winch. Taking from the injury prevention literature (Haddon, 1972), the most effective control of the hazard would be to “Prevent the creation of the hazard in the first place” by eliminating the use of the deck winch. This type of winch is not used in the European Union and some operators in the United States have eliminated it from their operation as well. However, if this cannot be done, engineering controls are the most desirable type of hazard intervention because they separate the worker from the hazard and decrease the possibility of an incident occurring.By using injury epidemiology to identify problems, along with practical industry input, effective safety interventions to control such hazards can be designed and implemented.
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