新型消声器用不锈钢管件加工专用设备设计含9张CAD图
新型消声器用不锈钢管件加工专用设备设计含9张CAD图,新型,消声,器用,不锈钢管,加工,专用设备,设计,cad
新型消声器用不锈钢管件加工专用设备设计 摘 要 新型消声器可有效消除噪声并有效降低汽车的尾气排放量,降低环境污染, 因此该产品具有重要的社会意义和经济价值,市场前景非常广阔。新型消声器用 不锈钢管件是新型消声器的主要零件,设计新型消声器用不锈钢管件加工专用设 备对于提高消声器的生产效率和产品质量都具有重要意义。本课题的主要任务是 设计一台用于加工新型消声器用不锈钢管件的专用设备,对不锈钢管件加工的各 种工艺和方法进行分析、比较,对日产量为 300 件的不锈钢管件加工 专用设备 进行总体设计,包括机械部分和电器部分,机械部分主要设计了减速器、丝杠、 导轨、气压系统、和机架,电气控制部分主要设计继电器-接触器控制系统的基 本控制电路。且设备的加工精度、生产率、可靠性等指标符合企业要求,设计的 装置应操作方便。设计过程中,减速器、丝杠的设计占主要部分,先对零部件进 行尺寸设计,然后对其进行必要的强度、刚度校验。液压系统的设计是先根据气 缸的载荷、行程等参数选取标准气缸产品和其他气压元件,然后根据工作机构运 动要求和结构要求进行气压系统图的设计。电器部分在本设备中主要是控制电机 的正反转。经过两个月的努力,以上工作已经完成。 关键字:减速器;涡轮蜗杆;丝杠;导轨;气压传动;继电器-接触器控制系统 Abstract New muffler can effectively eliminate the noise and exhaust of motor vehicles to reduce emissions, reduce pollution, so the product has an important social significance and economic value, the market prospects. New type of silencer used in stainless steel tube is a new muffler of the main parts, new muffler design uses stainless steel processing equipment muffler for improving production efficiency and product quality are of great significance. Stainless steel processing equipment can improve production efficiency and product quality . The main tasks of this task is to design a new type of muffler for the processing of stainless steel pipe fittings used in special equipment, stainless steel pipe fittings for a variety of processing techniques and methods of analysis, 300 of the output for the processing of stainless steel equipment for the overall design include mechanical parts and electrical parts. Mechanical parts of the main design include the reducer, screw, rail, air pressure systems, and rack. The relay - contactor control system is the basic control circuit. Machining accuracy, productivity, reliability indicators in line with corporate requirements, the device design easy to operate. Reducer, screw design accounted for the main part, first design the size of the parts, then the strength and rigidity is checked. Standard components of pneumatic system are designed. Partial control in the electric motor rotation。 Key words: Reducer; Turbine worm; Screw; Guide; Pressure transmission;Relay - contactor control system 目 录 1 传动机构的拟定 .6 1.1 机械传动参考方案 .6 1.1.1 方案比较.7 1.1.2 确定设计方案.8 1.2 流体压力传动参考方案 .8 1.2.1 液压传动的特点.8 1.2.2 气压传动的特点.8 1.2.3 确定设计方案.9 2 电动机的选择 .9 2.1 选择电动机的系列 .9 2.2 计算电动机功率 .9 2.3 确定电动机转速 .10 2. 4 选择电动机型号 .11 3 传动系统的运动学和动力学的计算 .11 3.1 计算总传动比和各级传动比的分配 .11 3.1.1 计算总传动比.11 3.1.2 各级传动比的分配.11 3.2 计算传动装置的运动和动力参数 .11 4 零件的设计计算 .13 4.1 传动零件的设计计算 .13 4.1.1 蜗杆蜗轮的设计.13 4.1.2 蜗轮与蜗杆的主要参数和几何尺寸.14 4.1.3 蜗杆传动的受力分析.16 4.1.4 校核齿面接触疲劳强度.16 4.1.5 校核齿根弯曲疲劳强度.17 4.1.6 校核蜗杆的刚度.17 4.1.7 精度等级公差和表面粗糙度的确定.18 4.1.8 蜗轮蜗杆的机构设计.18 4.2 轴的设计计算及校核 .19 4.2.1 蜗轮轴的设计.19 4.2.2 蜗杆轴的设计.23 4.3 轴承的选择和计算 .27 4.3.1 蜗轮轴上轴承的选择和计算.27 4.3.2 蜗杆轴上轴承的选择和计算.29 4.4 键连接的选择和校核 .30 4.4.1 蜗轮轴上键的选择和校核.30 4.4.2 蜗杆轴上键的选择和校核.31 4.5 联轴器的选择和校核 .31 4.5.1 蜗轮轴上联轴器的选择和校核.31 4.5.2 蜗杆轴上联轴器的选择和校核.32 4.6 箱体的设计 .32 4.6.1 传动件及箱体轴承座位置的确定.32 4.6.2 箱体结构形式和材料的选择.32 4.6.3 箱体的主要结构和尺寸关系.33 4.6.4 箱体结构的结构性和工艺性.33 5 润滑和密封的选择和计算 .34 5.1 润滑的选择和计算 .34 5.1.1 齿轮的润滑.34 5.1.2 减速器中传动件通常用浸油润滑.34 5.1.3 轴承的润滑.34 5.2 密封的选择 .35 6 减速器附件的选择 .35 6.1 通气器 .35 6.2 轴承盖(材料为 HT150) .35 6.3 油面指示器 .35 6.4 油塞 .35 6.5 窥视孔及视孔盖 .36 6.6 起盖螺钉 .36 6.7 定位销 .36 7 滑动丝杠的设计 .36 7.1 滑动丝杠螺母机构的传动方式 .36 7.2 滑动螺母结构 .36 7.3 丝杠支承 .37 7.4 丝杠牙形的选择 .37 7.5 螺距的选择 .37 7.6 丝杠直径的确定 .37 7.7 螺母长度 H 的确定 .39 7.8 丝杠螺纹部分 的长度 .39L螺 纹 8 滑动导轨 .40 8.1 导轨材料的要求和搭配 .40 8.2 导轨的结构 .41 8.3 滑动导轨的验算 .41 9 液压系统的设计 .44 9.1 气缸的选择 .44 9.2 气压原理图 .46 10 电气原理图 .47 11 设备执行机构动作说明 .48 结论 .48 参考文献 .49 谢 辞 .50 引 言 随着我国经济和工业的发展,噪声污染和空气污染已经严重影响人们的日常 生活,在大城市中噪声污染和空气污染主要来自于汽车尾气排放,这就需要一种 可有效消除噪声并有效降低汽车的尾气排放量的新型消声器,并能得到广泛使用。 为此本课题主要是设计一种能够自动加工新型消声器用不锈钢管件的设备。因为 本设备主要是切割管件,机构动作比较简单,采用数控控制代价较高。所以本设 备在传统机械传动的基础上添加气压传动,并通过简单的继电器-接触器控制系 统对电机的控制,实现自动加工的目的。 1 传动机构的拟定 1.1 机械传动参考方案 方案 1 方案 2 方案 3 方案 4 1.1.1 方案比较 方案 优点 缺点 外带式单级圆 柱齿轮减速器 (方案一) 结构简单,价格便宜。传递效率 中上,工作平稳性较好,有过载 保护,要求制造及安装精度较低, 润滑要求不高,环境适应性一般。 小功率传动,单级传动比较小, 外轮廓尺寸较大,传动精度低, 无自锁能力,使用寿命短,缓冲 吸振能力不好。 两级展开式圆 柱齿轮减速器 (方案二) 大功率传动,传动效率高,单级 传动比适中,传动精度高,使用 寿命长,环境适应性一般。在减 速器中应用最广泛,常用于载荷 较平稳的场合。 工作平稳性一般,缓冲吸振能力 差,无过载保护,要求制造及安 装精度高,无自锁能力,润滑要 求高。齿轮相对于轴承不对称分 布,要求级具有较大刚度。高速 轴应布置在远离扭矩输入端的一 边,以减小弯曲变形而引起的载 荷沿齿宽分布不均的现象。 两级同轴式圆 柱齿轮减速器 (方案三) 大功率传动,传动效率高,单级 传动比适中,传动精度高,使用 寿命长,环境适应性一般。箱体 长度较小,两大齿轮浸油深度可 大致相同。 轴向尺寸及重量较大,高速级齿 轮的承载能力不能充分利用,中 间轴较长,刚度差,仅能有一个 输入端和输出端,限制了传动布 置的灵活性。 单级蜗杆减速 器 (方案四) 传动比大,结构简凑,外输入与 输出轴垂直交错传动,冲击载荷 小,传动平稳,可有自锁能力, 传动精度高,价格相对便宜。适 合在工作温度较高、潮湿、多粉 尘、易爆、易燃场合适用。下置 式蜗杆减速器润滑条件较好,应 优先选用(v 4m/s ) 。 传动效率低,仅适用于中小功率 传动。无过载保护,制造及安装 精度要求高,要求润滑条件高。 当蜗杆圆周速度太高时(v4m/s) , 搅油损失大,采用上置式,此时, 蜗轮轮齿浸油、蜗杆轴承润滑差。 1.1.2 确定设计方案 本课题由指导老师指定使用单级蜗杆减速器这种传动方案,我分析其原因为 如下:由工作条件可知,载荷要求平稳,室内有粉尘环境下,应差用结构简凑, 传动平稳,且对工作环境要求不高的减速器,而单级蜗杆减速器(闭式)是较佳 选择。单级蜗杆减速器的使用寿命较长,能满足较长的大修期限和使用期限。 而对于可加工 7-8 级精度齿轮及蜗轮的中等规模机械厂,其 10 台批量生产的生产 力来说,生产能力满足其实际情况和要求,且生产单级蜗杆减速器相对成本较低, 盈利率较大。较大的一级传动比也是本课题首先考虑涡轮蜗杆减速器的主要原因。 综上所述,对于本设计要求,使用单级蜗杆减速器是较优设计方案,其各项特征 都满足设计要求,其整体性能优于其他方案。故,本文最终选用单级蜗杆减速器 作为该带式运输机的传动装置。 1.2 流体压力传动参考方案 1.2.1 液压传动的特点 (1)与电机比较,在同等体积下,液压装置能产生更大的动力,即具有大 的功率密度或力密度。 (2)液压装置容易做到对速度的无极调节,而且调速范围大,并且对速度 的调节还可以在工作过程中进行。 (3)液压装置工作平稳,换向冲击小,便于实现频繁换向。 (4)装置易于实现自动化,可以很方便地对液体的流动方向、压力和流量 进行调节和控制。 (5)由于液压传动中的泄露和液体的可压缩性使这种传动无法保证严格的 传动比。 (6)液压传动对油温的变化比较敏感,不宜在较高或较低的温度下工作。 1.2.2 气压传动的特点 (1)气压传动的工作介质是空气,它的粘度很低,所以流动阻力小,压力 损失小,便于集中供气和远距离输送。 (2)气压传动动作速度及反应快。液压油在管道中的流速一般在 5m 以下, 而气体流速可以大于 10m,甚至接近声速,可以在很短的时间内达到所要求的工 作压力及速度。 (3)气压传动有较好的自保持能力。 (4)气压传动系统工作压力低,仅仅适用于小功率场合。 1.2.3 确定设计方案 本设备中执行机构动作频繁,且又有一定的产量要求,所以就要求流体压力 传动动作速度及反应要快;另外要求设备不受外界条件变化的影响,对流体传动 的功率要求不高。通过对液压传动和气压传动特点的比较,最终选用气压传动。 2 电动机的选择 2.1 选择电动机的系列 按工作要求和条件可知,选取 Y 系列一般用途全封闭自扇冷鼠笼型三相交流 异步电动机,电压 380V。 2.2 计算电动机功率 电机 1 带动圆锯片转动,所以电机 1 的功率由圆锯片的切割功率确定。本设 备所用圆锯片为德国产圆锯片,其编号:720200320,外径: 200mm,齿厚:d外 =2.5mm,内径: 32mm,齿数:80。bd内 在切割管件的过程中,圆锯片需要一定的转速和进给速度,所以预先取圆锯 片转速为 = = ,进给速度为 = , 为圆锯片的一个720/minr6/rs4/msDA 齿所切割的面积: 22.50.160/80DbvAbr 切割力: 33.5cFKN 查金属工艺学表 1-2, 。MP 圆锯片的圆周线速度: 2017.40/dv ms外 外 线 切割力功率: 3310.Pkw线 查机械设计手册得: 滚动轴承效率: ;蜗杆传动效率: (双头) ;.9承 .8蜗 杆 联轴器效率: (弹性联轴器) ;滚筒效率:25联 0.96滚 筒 则电机 1 和圆锯片间的传动效率为: .0.9258承联 电机 1 所需功率: 10.74.19825Pkw电 电机 2 通过减速器带动丝杠转动,所以电机 2 的功率由丝杠确定。丝杠的功 率包括两部分,一个是克服管件对圆锯片的反作用力所消耗的功率 ,另一个是1P 克服导轨间摩擦力所消耗的功率。 在切割过程中,总切割力分解为两个力,一个是圆锯片切割管件的切割力, 一个是进给方向上的进给力。这里丝杠承受的是进给方向的进给力,而总切割力 的功率分配是圆锯片切割管件的切割力占总功率的 ,进给力功率占总809: 功率的 。:1 5 则 。1.5%0.19Pkw 因为圆锯片相对于管件的速度为 = ,从而得出管件对圆锯片的反作用1v4/ms 为: 10.30.FN 估计导轨上电机与支座总重为 ,查理论力学的出导轨间的摩擦系2kg ,则摩擦力的大小为:0.3.9.85f 丝杠推动电机 1 和支座运动,推力为: 10.358.0Ff N合 功率为: 331058.0.4.2PFvkw合 合 传动总效率 : 23. .2承 蜗 杆 联 丝 杠 所需电机功率: 2.k合电 机 由于本设备连续运转稳定,且传动效率较小,故只需使电动机的额定功率等 于或稍大于电动机实际输出功率就可以了。 2.3 确定电动机转速 预先设定丝杠螺距为 ,导程为 , 因为圆锯片相对于管件的速度为2m4 = ,既动导轨和支承导轨间的相对速度为 。所以丝杠的转速为:1v0.4/ms m2innr 根据机械设计基础中查得蜗杆的传动比在一般的动力传动中; i=740(常用值),最大值为 80。该传动方案为单级传动,则其相应电动机的转速 的范围应为: (740)12(7408)/minni r:电 2. 4 选择电动机型号 电机 1 是用来带动刀片转动的,必须具有一定的速度,由于额定功率相同的 同类型电动机,有四种常用同步转速,即 3000、1500、1000、750r/min。电动机 的转速越高,极对数越少,尺寸和质量就越小,价格也越低。电机 1 是用来带动 刀片转动的,对钢管进行切割,必须具有一定的速度和切割力,所以不能一味的 追求高速度;电动机 2 是用来带动丝杠转动的,若选用电机的转速太高,将使传 动装置的传动比越大,从而使传动装置的结构尺寸也跟着增大,整个减速器的成 本就越高。因此,对电动机及传动装置做整体考虑,综合分析比较,以上述算出 的电动机输出功率 Pd 和电动机转速范围 Nd 查机械设计手册,电机 1 选择 Y160M1-8,额定功率 ,满载转速 ,同步转速 ;电机 24kW750/minr720/minr 选择 Y802-2,额定功率 ,满载转速 ,同步转速 。1.385 3 传动系统的运动学和动力学的计算 3.1 计算总传动比和各级传动比的分配 3.1.1 计算总传动比 根据电动机满载转速 nm 及工作机转速 n,可得传动装置的总传动比: 2583.410ai 3.1.2 各级传动比的分配 由于为单级蜗杆传动,传动比都集中在蜗杆上,其他部分不分配传动比, 其传动比就为总传动比: 23.54ai 3.2 计算传动装置的运动和动力参数 电机轴: 02.05Pkw电 机 8/minnr 300.5109.5.728PTNmn 蜗杆所在轴: 10.9.46kw联 10n285/inr311.19.50.6PTNm 蜗轮所在轴: 210.496.08.392kw承 蜗 杆12851./min3.2nri3229.50.29.PTN 丝杠支承: 320.3.5.385kw承联 2198/minnr333 .109.555PT N 各轴运动及动力参数汇总如下表 2-1: 表 2-1 轴序号 功率 P/ kW转速 n/(r/min) 转矩 T/ Nm:传动形式 传动比 效率 电机轴 0.005 2825 0.017 蜗杆所在轴 0.00496 2825 0.016 蜗杆传动 23.5 0.80 蜗轮所在轴 0.00392 119.998 0.312 丝杠支承 0.00385 119.998 0.306 联轴器 1 0.9925 4 零件的设计计算 4.1 传动零件的设计计算 一般情况下,首先进行箱外传动件的设计计算,以便使减速器设计的原始条 件比较准确。在设计箱内传动件后,还可以修改箱外传动件尺寸,使传动装置的 设计更为合理。由于本方案为单级蜗杆传动,无箱外传动件,故直接进行箱内传 动件的设计。 4.1.1 蜗杆蜗轮的设计 蜗杆类型、精度等级及齿数的确定 根据本设计要求,宜选用圆柱蜗杆,而圆柱蜗杆中宜选用普通圆柱蜗杆。再 根据 GB/T10085-1988 的推荐,采用蜗杆传动的装置,宜选用渐开线蜗杆传动 (ZI 蜗杆) 和锥面包络蜗杆(ZK 蜗杆) 。由于 ZK 蜗杆一种非线性螺旋齿面蜗杆,不能 在车床上加工,只能在铣床上铣制并在磨床上磨削成品,其精度较高,造价较贵。 而渐开线蜗杆易于加工,成本较低,故考虑实际使用情况和成本等因素,选用 ZI 蜗杆为最优选择。 (机械设计239 页)由该厂生产条件及设计要求限制,且在 一般工业中应用 7-8 级精度的蜗杆,故本蜗杆选用 8 级精度。 (机械设计257 页)蜗杆头数由传动效率决定,取 z1=2。为了提高稳定性和传动效率,蜗轮选用 斜齿蜗轮。 蜗杆蜗轮材料的选择 考虑到蜗杆传递功率不大,转速也只是中等,要求寿命为 17280h(一班制 10 年)故蜗杆材料应选 20Cr,表面渗碳淬火处理,使齿面硬度 .蜗轮选56HRC 用耐磨性最好的铸锡磷青铜(ZCuSn10P1) ,金属模铸造。为了节约贵重的有色金 属,仅齿圈用青铜制造,而轮芯用灰铸铁 HT150 制造。为了防止变形,常对蜗轮 进行时效处理。 按齿面接触强度设计: 根据闭式蜗杆传动的设计准则,先按齿面接触疲劳强度进行设计,再校核齿 根弯曲疲劳强度。传动中心距: 232()EHZaKT (1)确定作用在蜗轮上的转矩 T2: 3220.9219.50.28PTNmn (2)确定载荷系数 :K 因工作载荷平稳,故载荷分布不均系数取 =1;由载荷均匀无冲击,查机K 械设计手册取使用系数 KA=1;由转速不高,动载荷系数应取 KV=1.1;故载荷系数 1.1 AVK (3)确定弹性影响系数 ZE: 因选用铸锡磷青铜(ZCuSn10P1)蜗轮和钢蜗杆配对,故 ; 1260EZMPa (4) 确定接触系数 : 先假定蜗杆分度圆直径 d1 和传动中心距 a 的比值 ,查机械设计1.35d 图 11-18 得, =2.9;Z (5)确定许用接触应力 :H 根据涡轮材料为 ZCuSn10P1,金属模铸造,蜗杆螺旋齿面硬度56HRC,查 机械设计表 11-7 得蜗轮的基本许用接触应力 。268HMPa 蜗杆的工作寿命: ,2130417hLh 蜗轮轮齿的应力循环次数: 769.01h 寿命系数: 778810.32HNKPa 蜗轮齿面的许用接触应力为 =0.730HNHK268195MPa (6)计算中心距得: 22332160.9()().13.05EHZaKTm 根据经验取中心距 a=200mm,因 i=23.54,故查机械设计表 11-2 中取 模数 m=8mm,蜗杆分度圆直径 d1=80mm.这时 d1/a=0.4,查机械设计图 11-18 得接触系数 28(机械设计244 页) ;模 数与蜗杆相同为 8mm;变位系数 。 20.5x 18afmm 24t 验算传动比: 10.5zi 这时传动比误差为: 是不允许的,查机3.54.%1.92 械设计表 11-1:当 i=1430 时,宜选 2961。 再次是试选 z2 =45,此时传动比误差为: 453.425 故最终选取 。2z 计算蜗轮的主要参数如下: 蜗轮分度圆直径 : 2845360dmzm 蜗轮喉圆直径: 2()8aah 蜗轮齿根圆直径: 2 1432ff 蜗轮咽喉母圆半径: 210gar 蜗轮宽度: 20.75.967bd 蜗轮齿顶圆弧半径: 1/3aRm 蜗轮螺旋角: = =111836 蜗杆传动的标准中心距为: 122()()1045820dqzm 蜗杆变位后中心距为: .6axm 4.1.3 蜗杆传动的受力分析 不计摩擦时,其各力计算如下: 11230.6.48taTFNd 21 3.1.7at 12tan.tan20.6.43rFN 232 1.82cos1.9.0nnTd 4.1.4 校核齿面接触疲劳强度 原始接触应力公式为 0nHEKFZL: 由上述已算出数据知,啮合面上的法向载荷 ,载荷系数 K=1.1,1.82nN 材料的弹性影响系数 。 126EZMPa 验算公式为: 3HEHKTZ: 由上述已算出数据知,蜗轮公称转矩 ;中心距 a=216mm;载20.1Nm: 荷系数 K=1.1;材料的弹性影响系数 ;蜗杆分度圆直径 d1 和传动中 1260EZMPa 心距 a 的比值 ,查机械设计图 11-18 得, =2.7。10.37dZ 231.3602.725196HE HTaMPaa : 满足蜗轮的齿面接触疲劳强度要求。 4.1.5 校核齿根弯曲疲劳强度 21.53FFaFKTYdm 当量齿数 : 23345.92cos(.6)vz 根据 , ,从机械设计 图 11-19 中可查得齿形系数20.5x24.9vz 。.8FaY 螺旋角系数: 1.3610.904Y 许用弯曲应力: FHNK: 查机械设计表 11-8 宏 ZcuSn10P1 制造的蜗轮的许用基本弯曲应力为: MPa56F 寿命系数: 69710.582.4HN MPa.93.20F .531.6806 FMPa 故弯曲强度是满足的。 4.1.6 校核蜗杆的刚度 刚度校核公式为: 21348trFyLyEI 由上述已算数据知,蜗杆所受的圆周力 ;蜗杆所受的径向力10.4tN ;蜗杆材料的弹性模量 E=206000MPa;蜗杆的危险截面的惯性矩10.643rFN , 441 5463.510.8fdI m 一般情况下, ,蜗杆两端支承间的跨距 。许用最2.9.3602Ld 大挠度 则:10.dym 22399.40680.1.58081.y m 满足蜗杆的刚度要求。 4.1.7 精度等级公差和表面粗糙度的确定 考虑到所设计的蜗杆传动是动力传动,属于通用机械减速器,从 GB/T10089- 198 圆柱蜗杆、蜗轮精度中选择 8 级精度,侧隙种类为 c(115 ) ,蜗杆的标注m 为 8 c GB/T10089-1988,蜗轮的标注为 7-8-8 f GB/T10089-1988。然后由机 械设计手册查得要求的公差项目及表面粗糙度如下: 公差配合为:H7/s6 GB/T1800-79 形状和位置公差:查国标 GB11821184-80 表面粗糙度:Ra=1.6 GB1031-83m 4.1.8 蜗轮蜗杆的机构设计 蜗杆和轴做成一体,即蜗杆轴。为保证刚度,应采用无推倒槽结构 图 3.2 蜗轮采用齿圈式(如下图 3.3) ,青铜轮缘与铸造铁心采用 H7/s6 配合,并加 台肩和螺钉固定,螺钉选 6 个。螺钉拧入深度为 0.30.4B,即 21.628.8mm。 图 3.3 4.2 轴的设计计算及校核 轴是组成机器的主要零件,一切作回转运动的传动零件,都必须安装在轴上 才能进行运动及动力传递。 4.2.1 蜗轮轴的设计 (1)轴类型、材料的选择 按承受载荷的不同,轴可分为转轴、心轴和传动轴。按轴线形状不同,分为 曲轴和直轴,直轴又根据外形不同,分为光轴和阶梯轴。本设计方案的工作轴既 承受扭矩,又承受弯矩,且要求零件装配定位要精确,故本方案选择阶梯转轴。 轴的材料主要是碳钢和合金钢。根据本设计要求,对轴强度、刚度等方面的要求 不是很高,碳钢价格低廉,对应力集中敏感度低,同时可用热处理或化学热处理 提高其耐磨性和抗疲劳强度,故本方案选用 45 钢,调质处理。 (2)蜗轮轴的功率、转速和转矩 210.496.08.392Pkw承 蜗 杆85/min3.nri 322.219.54TN (3)作用在蜗轮上的力 284360dmzm 112.4taTFN 21 30.1.76atd 12tan.tan2.60.43rF 各力方向如图 3.1 所示。 (4)轴的最小直径的初步确定 查机械设计表 15-3,取 ,于是得:05A 33min.421.689Pdm 蜗轮轴的最小直径显然是安装联轴器处的直径 d- (如下图 3.5) 为了使所选的轴直径与联轴器孔径相适应,故需与联轴器型号同时确定。 联轴器的计算转矩 ,查机械设计表 14-1,考虑到转矩变化很小,故2caATK 取 KA=1.5,则 mm21.503468ca N 按照计算转矩 应小于联轴器公称转矩的条件,查国标 GB4323-84,选用 HL5 型弹性柱销联轴器,其公称转矩的条件为 1000000Nmm。本联轴器的孔径 d1=60mm,故 d- =60mm,半联轴器长度为 L=142mm,半联轴器与轴配合的毂孔长 度 L1=84mm。 (5)轴的结构设计 拟定轴上零件的装配方案 图 3.5 根据轴向定位的要求确定轴的各段直径和长度 为了满足半联轴器的轴向定位要求,-轴段右端需制出作为轴肩,故取 -段的直径 d- =66mm,左端用轴端挡圈定位,按轴段直径取挡圈直径 D=62mm。半联轴器与轴配合的轂孔长度 L1=107mm,为了保证轴端挡圈只压在 半联轴器上而不压在轴的断面上,故-段的长度应比 L1 略短一些,现取 l- =104mm。 初步选择滚动轴承。因轴承同时受有径向力和轴向力的作用,故选用单列 圆锥滚子轴承。参照工作要求并根据 d- =66mm,从轴承产品目录中初步选取单 列圆锥滚子轴承 7214E,其尺寸为 ,故取 d- =701256DTm d- =70mm,轴套尺寸为 D=72mm,l=16mm,而 l- =26+16=42mm。右端滚动 轴承采用轴肩进行定位,有机械设计手册查得 TL9 型轴承的定位轴肩高度 h=3mm,因此,取 d- =74mm。 取安装蜗轮处的轴段-的直径为 d- =74mm;蜗轮的左端与做轴承间 采用套筒定位。蜗轮轮毂宽度为 B=72mm,为了使套筒断面可靠地压紧蜗轮,此 轴段应略短于轮毂宽度,故取 l- =108mm。蜗轮的右端采用轴肩固定,轴肩高 度 h(0.07 0.1)d,故 h=3mm,则轴环处的直径 d- =78mm。轴环宽度: =54mm,双轴肩,用以保证蜗轮位置的对中性。取 d- =70mm, l-1.4b =12mm, L- =26mm。 轴承端盖的总宽度为 20mm(由减速器及轴承端盖的结构设计而定) 。根据 轴承端盖的装拆及便于对轴承添加润滑脂的要求,取端盖的外端面与半联轴器的 左端面间的距离为 l=30mm,故取 l- =50mm。 至此,本蜗轮轴的各段直径和长度基本确定下来,其数据汇总如下表: 表 3-1 尺寸 - - - - - - 直径 (mm) 60 66 70 74 78 70 长度 (mm) 104 50 42 108 12 26 轴向零件的周向定位 蜗轮、半联轴器与轴的周向定位均采用平键连接。按 d- 查机械设计 表 6-1 得平键截面 ,键槽用键槽铣刀加工,长为 63mm,同201bhm 时为了保证蜗轮与轴配合有良好的对中性,故选择蜗轮轮毂与轴的配合为 ;76Hn 同样,半联轴器与轴的连接,选用平键为 ,半联轴器与轴的810m 配合为 。滚动轴承与轴的周向定位是由过渡配合来保证的,此处选轴的直径76Hk 尺寸公差为 m6。 轴上圆角和倒角尺寸的确定 参考机械设计表 15-2,取轴端倒角为 ,各轴肩处的圆角半径如图245 4.5,一般取 R=2mm。 轴上载荷的计算 首先根据轴的结构图(图 3.5)做出轴的计算简图(如下图 3.6) 。查机械设 计手册,对于 TL9 型圆锥滚子轴承,取 a=21mm,从而确定轴承的支点位置。由 图 3.5 可得,作为简支梁的轴的支承跨距: L1=94mm,L 2=94mm, .29418m 再根据轴的计算简图做出轴的弯矩图和扭矩图。 图 3-6 从轴的结构图以及弯矩图和扭矩图中可以看出截面 C 是轴的危险截面。现计 算出的截面 C 处的 MH、 MV 及 M 的值列入下表 3-2: 表 3-2 载荷 水平面 垂直面 V 支反力 F0.867,0.867NHaNHbF0.32,0.32NVaNbF 弯矩 12m141mM 总弯矩 22.67M 20609 扭矩 T0.31TN 按弯扭合成应力校核轴的强度 进行校核时,通常只校核轴上承受最大弯矩和扭矩的截面(即危险截面 C) 的强度。根据上表 4.2 中的数据,以及轴单向旋转,扭转切应力为脉动循环变应 力,取 ,则轴的计算应力为:6.0 2 2220 63.190.61.591074ca aMT MPW 前已选定轴的材料为 45 钢,调质处理,查表可得 。因此 ,a11c 故安全。 4.2.2 蜗杆轴的设计 (1)轴类型、材料的选择 蜗杆轴类型同蜗轮轴类型,都为阶梯转轴。蜗杆轴材料由定蜗杆材料时确定, 故本轴材料为 45 钢。 (2)蜗杆轴的功率、转速和转矩 10.50.92.46Pkw联 1n8/minr311.19.5.0TN (3)作用在蜗轮上的力 d1=80mm 11230.6.48taTFNd21 3.1.7at 12n1.7tan0.6.43rtF 各力方向如图 4.1 所示。 (4)轴的最小直径的初步确定 查机械设计表 15-3,取 A0=112,于是得: 3233min.921015.8Pd m 蜗轮 Deformation and failure mechanisms of lattice cylindrical shells under axial loadingYihui Zhang, Zhenyu Xue, Liming Chen, Daining Fang *Department of Engineering Mechanics, FML, Tsinghua University, Beijing 100084, PR China1. IntroductionThe interest of lattice structures with various core topologies has grown rapidly over the last decade for their superior properties of high specic stiffness and strength, effective energy absorption, shock mitigation and heat insulation 14. These studies have shed light on that well-designed lattice structures are able to outperform the solid plate and shell components in many applications. In general, the structural topology, as the primary concern in the design, plays a signicant role in dominating the overall mechanical response of the structures. Understanding the deformation mechanisms of various topologies undoubtedly aids to attain the best design. Most of the previous studies were focused on the twodimensional planar lattices. Figs. 1(a)(d) exhibit four types of the planar lattice congurations, namely diagonal square, hexagonal, Kagome and triangular, respectively. Each of them has the periodic patterning formed from a two-dimensional geometric shape with an innite out-of-plane thickness. The overall effective in-plane stiffness and strength of the diagonal square, hexagonal, Kagome and triangular lattices have been analyzed recently, and they show a rich diversity in deformation 58. For the diagonal square and hexagonal lattice plates, each truss member undergoes bending deformation under most in-plane loading conditions, except for the diagonal square lattice plate uniaxially loaded along the axial directions of its truss member. For the triangular and Kagome lattice plates, the deformations of their truss members are always dominated by their axial stretching or compressing, resulting in higher stiffness and load capacity than the former two. The hexagonal lattice structure can be processed easily using standard sheet metal fabrication method. The elastic modulus, plastic yield as well as buckling behavior of the hexagonal honeycomb have been extensively explored 1,9,10. A new kind of fabrication method named powder processing technology has been developed recently 11, thus activating more varieties of complicated congurations to be fabricated by this approach. Wang and McDowell 8,12 systematically analyzed the stiffness, strength and yield surfaces of several types of planar lattice patterns. Fleck and Qiu 13 estimated the fracture toughness of elastic-brittle planar lattices using nite element method for three topologies: the hexagonal, triangular and Kagome lattices. Zhang et al. 14,15 proposed two novel statically indeterminate planar lattice structures and furthermore formulated their initial yield surfaces and utmost yielding surfaces. As an ultra-light-weight material, lattice material is an ideal candidate of traditional material in aerospace engineering. For example, utilizing the winding technology, one can manufacture lattice cylindrical shells, which, as depicted in Figs. 1(e)(h), are the key components of aerospace craft and airplane. The three dominating geometrical parameters of the representative unit cell are demonstrated in Fig. 2 by exemplifying the triangular lattice cylindrical shell, where is the arc Fig. 1. Congurations of four 2D lattice plates and the corresponding cylindrical shells: (a) diagonal square lattice plate; (b) hexagonal lattice plate; (c) Kagome lattice plate; (d) triangular lattice plate; (e) diagonal square lattice cylindrical shell; (f) hexagonal lattice cylindrical shell; (g) Kagome lattice cylindrical shell; (h) triangular latticecylindrical shell. length of each beam,and denote the thickness of the beam in the radial direction of the cylinder and the thickness of the beam in the shell face, respectively. The hexagonal lattice sandwich cylindrical shell has been popularly utilized in practical applications as fuselage section of aircrafts and load-barring tubes of satellites for several decades 1619. Under axial compression, this lattice sandwich cylindrical shell possesses better mechanical performance than the traditional axial stiffened cylindrical shells. Although much attention has been paid on the mechanical behavior of hexagonal lattice, the previous investigations were mainly focused on the simple planar hexagonal lattice structures such as beams and plates, and the delicate investigation on the mechanical behavior of hexagonal lattice cylindrical shell has been scarce. Therefore, the lattice cylindrical shell of hexagonal topology is one focus in this paper. The lattice cylindrical shells made from Fig. 2. The sketch of triangular lattice cylindrical shell with one unit cell in the axial direction (a) and the three-dimensional gure of the representative unit cell with illustration of its geometric parameters (b).Fig. 3. Deformation modes and the non-dimensional axial elastic moduli of hexagonal lattice cylindrical shells with the beam of rectangular cross section and the geometric parameters,; ; stretching- dominated topologies, such as the triangular and Kagome lattices, due to their better in-plane mechanical property than that of the hexagonal one 1,8, are likely to be better candidates to the axial stiffened cylindrical shell than the hexagonal lattice cylindrical shell. Under operating, the cylindrical shell must be able to bear relatively large axial load and resist buckling.Sophisticated analyses of the axial elastic modulus, the axial yield strength and the axial bulking of the lattice cylindrical shell is crucial when assessing its performance under axial loading. Noting that the studies are still comparatively scarce for the mechanical behavior of these lattice cylindrical shells though there are some limited experiments results reported 20, the study on deformation mechanisms and failure analysis of lattice cylindrical shells should be valuable and benecial for practical applications of lattice structures in engineering. The outline of the paper is as follows. Section 2 focuses on the type of cylindrical shells made from the bending-dominated planar lattices. The emphasis is placed on the inuence of the geometric dimensions on the overall effective stiffness and elastic buckling behavior. In Section 3, we rst present simple models capable of quantitatively predicting the effective elastic modulus and yield strength of Kagome and triangular lattice cylindrical shells. The models are veried by the corresponding nite element calculations. Furthermore, we explore the failure modes of the Kagome lattice cylindrical shell and construct a failure mechanism map to identify them. Finally, a comprehensive comparison of their load capacities versus their weights is made among three types of lattice cylindrical shells, indicating that Kagome and triangular lattice cylindrical shells have similar load capacities and both outperform the hexagonal one.2. Axial mechanical properties of cylindrical shells made from the bending-dom-Inated planar latticesIn this section, the deformation mechanisms of cylindrical shells made from bending-dominated planar lattices are studied for the case of an axial stress uniformly distributed through the whole cylindrical shell. Two topologies of lattices, diagonal square and hexagonal, are considered as sketched in Figs. 1(e) and (f). The beams of the lattice structures are all ideally welded with each other. A static analysis based on nite element method has been carried out to identify the effective elastic modulus as a function of the geometric and material parameters. An analytical solution for the critical elastic buckling load is also provided utilizing the homogenization method. Because of the similarity of the deformation mechanism between the diagonal square and hexagonal lattices, only the result of the hexagonal topology is selected for demonstration.Finite element analyses using the commercial software ABAQUS are carried out for the hexagonal lattice cylindrical shells. Both circular and rectangular shapes are considered as the beam sections. The aspect ratio of the rectangular section, ,is dened as . The parent material of the lattice structure is isotropic, with the material parameters xed as and , where and denote Youngs modulus and Poisson ratio of the material, respectively. Beam elements (B32 in ABAQUS denotation) and rened meshes (30 elements for each beam) are adopted to ensure accuracy. The numerical results for the case of rectangular beam cross-section are summarized in Fig. 3. The axial elastic modulus is dened according to the average axial strain of the cylindrical shell under a given uniform axial stress. In FE calculations, a uniform axial stress is applied and the measured average axial displacement of the free end is used to estimate the effective axial strain. The effective axial modulus of lattice cylindrical shell is just calculated using the uniform axial stress and effective axial strain. As to the axial yield strength discussed in Section 3, it is measured based on the maximum stress within the lattice members. The yield of the cylindrical shell is assumed to be indexed by the yield of the lattice member. It is observed that the cross-section of the cylinders does not remain circular if their beam members have non-square cross-section in the elastic deformation range. While for a hexagonal cylindrical shell with beam members of a circular cross-section, its cylindrical crosssection keeps circular as the structure is deforming. The numeric results on the deformation mechanism can be summarized as that the cross-section of axially loaded hexagonal cylindrical shells will not remain circular if the inertia principal direction of the beam section is not arbitrary. It should be stated that this phenomenon may not be inconsistent with its application since this lattice cylindrical shell is commonly utilized as the core of sandwich structure and the rigidity of the two panels constrains the local bending to a great extent. Based on the nite element calculations, the normalized effective axial elastic modulus of the hexagonal cylindrical shells with a rectangular cross-section of beam member is also plotted in Fig. 3 as a function of the aspect ratio a. The relative density of the lattice cylindrical shell, , is fixed as . The number of unit cells in the axial and circumferential directions of the lattice cylindrical shells, and , are also kept as,and only the thickness in the radial direction of the shell, b, is changed. We dene the relative thickness of the cylindrical shell, , as the ratio of the thickness of the cylindrical shell to the radius, i.e. . Simple analysis gives that the effective modulus of the bending-dominated hexagonal planar lattice structure, can be expressed as 1. then is employed as the reference to normalize the calculated effective elastic modulus of the lattice cylindrical shells. It is found that the normalized elastic moduli of the lattice cylindrical shells is linearly dependent on the relative thickness approximately expressed by the common equation, , where denotes the axial effective elastic modulus of the lattice cylindrical shell. When the beam of cylindrical shell has a square cross-section, its axial modulus approximately equals to that of the planar counterpart, otherwise the axial modulus monotonously increases as the aspect ratio of the beam section increases. The qualitative explanation on their underlying deformation mechanisms is given as follows: the cross-section of the hexagonal cylindrical shell with the beams having square cross-section remains circular under axial loading. In this case, the axial deformation of the cylindrical shell is mainly contributed by the bending of the curved beams in the shell face, which is approximately the same as the bending of the beams in its planar counterpart under the same loading condition. Therefore, the effective axial elastic modulus of the hexagonal cylindrical shell should be approximately equal to its planar counterpart correspondingly. When increasing the aspect ratio, , and xing the product, bt, the bending deformation out of the shell face of the curved beam decreases, resulting in that the variation of the shell radius as well as the circumferential deformation decrease. Furthermore, the axial deformation decreases due to the Poisson effect. Hence, the axial effective modulus monotonously increases as the aspect ratio of the beam section increases.Elastic buckling of the hexagonal cylindrical shells is investigated both numerically and analytically. In the nite element calculations, the attention is rst restricted to the role of aspect ratio, . The geometric parameters are ,and in FE calculations, where the product of the thickness and is xed in order to x the weight of the lattice shell. Fig. 4 shows that the critical buckling load will be varied dramatically with the change of the aspect ratio in the range roughly between 0 and 4. Its value attains the maximum when. Akin to the fact that the initial curvature of a straight beam can signicantly reduce its critical buckling load, the nonuniform deformation of the lattice cylindrical shell that we described previously introducing an imperfection into its deformed conguration results in the reduction of the critical buckling load for the cases of Fig. 4. The buckling mode and the relationship of the non-dimensional overall elastic buckling load with the aspect ratio of rectangular beam section for the hexagon cylindrical shell with the geometric parameters and the Poisson ratio of the parent material, ,. The product of the thickness and is xed in order to x the weight of the lattice shell.To derive the solution for the critical buckling load when , we employ a homogenization approach such that the discrete lattice cylindrical shell is smeared out as a homogeneous solid shell of the same dimensions whose effective properties mimic those of the lattice one. For the homogeneous solid cylindrical shell, the critical buckling load can be written aswhere and are the Youngs modulus and Poisson ratio of the effective solid material, is the thickness of the shell and is a knockdown factor that multiplies the classical critical load for a perfect cylindrical shell. The overall buckling of a cylindrical shell is generally imperfection-sensitive and takes this into account. The empirical choice for as a function of is suggested by NASA 21,22 such that where This dependence of on is based on a lower bound to reams of critical load data on cylindrical shell buckling. The microstructure of the lattice cylindrical shell causes the diversity of the buckling modes. The geometry of the lattice microstructure determines the eigenvalues for different buckling modes, and therefore can change the coincidence of eigenmodes. So it is responsible for the extreme imperfection sensitivity of lattice cylindrical shells. In this paper, the NASA critical load knockdown factor is adopted as an approximation to consider the inuence of imperfection sensitivity. Both the analytical results with and without this knockdown factor, , are calculated for comparison. The effective elastic modulus and Poisson ratio of the hexagonal lattice cylindrical shell are approximately the same as that of its unwound counterpart. By assuming that the overall structural deformation is dominated by bending of each beam member and ignoring the contribution of its axial deformation, previous studies 23 provided the expression of the effective elastic modulus, Eplanar, and Poisson ratio, nplanar, of the planar hexagonal lattice as ,However, the above equation is not suitable for evaluating the critical buckling load since it combining with Eq. (1) predicts an innite value, which is meaningless. The more rigorous derivation of the effective elastic modulus, Eplanar, and Poisson ratio, nplanar,of the planar hexagonal lattice is performed by considering both the contribution of the bending and axial deformation of each beam. The detailed derivation can be found in the monograph of Gibson and Ashby 1, and here only the results are given as ,By comparison, Eq. (4) overestimates both the effective elastic modulus and Poisson ratio. The analytical solution of critical load for the hexagonal lattice cylindrical shell, , can be available by substituting Eq. (5) into (1), i.e. Fig. 5 illustrates the relationships of the non-dimensional critical buckling load with the number of unit cells in the circumferential direction at different weight levels. The non-dimensional critical buckling load and weight index of the lattice cylindrical Fig. 5. The non-dimensional overall elastic buckling loads of the hexagon cylindrical shell versus the circumferential numbers of unit cells at different weight levels with the beam of square cross section and the parameters, ,Fig. 6. The comparison of the analytical predictions and FE results for the nondimensional overall elastic buckling loads of the hexagon cylindrical shell at different relative densities with the beam of square cross-section.shell are dened as and , , respectively, where denotes the density of the material, is the height of the cylindrical shell and is the weight of the cylindrical shell, i.e. . The ratio of the height of the cylindrical shell to its diameter, , is kept constant as . Both analytical and numerical predictions are plotted. It is shown that at the same weight level, the analytical critical load incorporating the knockdown factor decreases gradually to a steady value with the increase of the circumferential number of unit cells. The prediction based on the analytical results without knockdown factor overestimates the buckling load by about 50%. Discrepancy between the numerical results and the analytical solutions with knockdown factor displays when the circumferential number of unit cells, , is small, e.g. . While the results from two methods are more consistent when n are larger than about 8. This is mainly due to the difference of loading conditions of the two methods such that the loads are discretely applied at limited points in FEM, while uniformly applied in the analytical homogenization method. The circumferential numbers of unit cells of the lattice cylindrical shells are usually larger than 8 in practical applications, so the proposed analytical model (Eq. (6) provides a good approximation of the critical elastic buckling load for the lattice cylindrical shells. Based on the FE results, there exists a maximum value in the range of n between 10 and 15 for each of the three different weight levels. Additional calculations are performed for the hexagonal cylindrical shells of equaling 10 and 15, respectively. For such two specied cases, the analytical predictions of the critical buckling loads varying with the relative density are compared with the corresponding nite element results in Fig. 6. Good agreement is found between the FE results and the analytical solution with the knockdown factor. Since the present analytical solution with the knockdown factor can predict the inuence of both two geometric parameters with high accuracy, as shown in Figs. 5 and 6, the present solution (Eq. (6) is suggested in practical applications. It should be noted that no initial geometrical imperfection is introduced in the nite element model of the lattice cylindrical shell. But as is different from the solid shell, the lattice shell is made of discrete curved beam, resulting in the high discreteness of the shell structure. This kind of discreteness of the cylindrical shell can be deemed as a kind of imperfection. The fact that the analytical model incorporating the knockdown fac
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