限位板的冲压模具设计【说明书+CAD】
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小型柴油机高速化和大功率化的发展DBroomeRicardo & Co.Engineers(1927) Ltd.(England) 笔者所在公司长期关注小型高速柴油机发展,特别是其燃烧系统的发展。仅管这类机型在北美大陆并不普及,但在欧洲和日本的产量以及使用量却相当大,约有数百万台。典型的机型是自然吸气四缸柴油机,每缸排量为25-35 in3 (400-600cm3),工作转速为4000-5000 rpm,活塞限速在2400ft/min(12m/s)。在warren的美国军方坦克研发中心(USATAC),Mich提议开发一种符合军方要求的大功率轻型动力装置,并且要求能达到前所未有的转速以实现性能的提升,而不是仅仅增加涡轮增压器。在这个情况下,进气系统和燃烧系统的研究就被搬上了前台。这项计划制定了工作的执行方案,包括设计,制造和在特定单缸机上的初步实验。项目这个项目的技术要求最后由USATAC制定,其在大概的内容如下所述:1 设计,采办,制造和测试一台缸径为3.5英寸(88.9mm)的单缸机。要求以最高转速工作,至少5000 rpm。以分离的空气供给系统的对涡轮增压状况进行模拟。2 改进这台单缸试验机,以实现其预定的性能指标,使得相对应的军用四缸机能够产生 1 bhp/in3 (45.5kW/cm3)的功率以实现单位质量功率3.5 lb/bhp (2.13kg/kW)。3 设计将不受传统观念影响,以最小化机械负荷和热负荷为目标。4 最初的燃料要求以CITE-R fuel (MIL-F-46005A (MR)为标准。一开始,专门针对航空汽油进行研究,但后来这项要求有所放宽。5 如果可能的话使用MIL-L-2104B规格的润滑油。6 项目的最后阶段包括一台军用四缸机的设计,其中包含了对单缸机的测试。7 对于多缸机的启动,怠速和小负荷工况的工作性能必须不能被忽视。初步设计方案一个对缸径尺寸和功率输出指标的试验可以很快地显示出引擎最高转速的极限。从最小转速5000 rpm开始,曲线清楚地显示出转速在6000 及以上时,这些活塞的运动速度与赛车引擎的相当。通过增加转速来减小平均有效压力的意义是重大的,fmep(这些估计都是根据笔者公司的一些过去的数据的分析,其中大部分被概括在图1)的增加对指示平均有效压力的减小几乎没有反馈。几年来,自然吸气柴油机被要求严格工作在冒烟界限以内,因此其指示平均有效压力只能提高到145lb/in2 (1000kPa);因此,涡轮增压的测量方面有了一些要求。高速和高引擎磨擦带来的恶果就是燃料的消耗。表格1中清楚地显示了bsfc对曲线的迅速恶化,效果并不比汽油机来的好,所以会损失压缩做功循环的主要利益。根据这种情况,就计划将转速限定在6000 rpm。主要的表现问题在引擎的设计中所要求的可以被概括为以下几项:引擎的进排气 根据以往在高速小型柴油机上的经验表明,进排气性能是活塞高速化时限制指示平均有效压力的重要因素。因此,需要提供足够大的充气系数使得活塞速度能达到3500 ft/min (17.8m/s),比现有的柴油机水平高出50%。这就必然要求设计脱离传统的气缸设计,包括更倾向于多气门的设计(表2);涡轮增压带来微小的变化,比如更高的进气温度,最小的压力损失和减小体积功率的变化。除此之外,涡轮增压匹配要求需要被事先确定。对于车用引擎,由于对后备扭矩的需要,通常倾向于在额定的转速前提下,最小化有效的推进力。因此大尺寸的排气门并不被强求。另外,绝对短的废气排放系统要求排气速度尽可能低,排气门面积与进气门相当。同时需要考虑的是气门开启时间的设置。在进气口,高速度通常联系着较大的气门关闭延时角,在柴油机里,一般为提前下止点45度,这将会连续地影响起动性能,同时也牺牲一定的低速性能,更大程度上影响到引擎的自然扭矩储备。对于排气,因为涡轮增压的需要,排气门在做功冲程结束前就要开启,在高转速下提前60度开启并没有什么优势。较大的气门重叠角可以减少废气量和降低排气系统各组成部分的温度,同时可以最小化高转速时充气系数的降低程度。充气系数与气体的流动有关。而由于活塞和气门间余隙容积产生的机械问题将会对燃烧系统有负面的影响。燃烧问题当引擎达到预定转速时,点火延时期的持续时间将是一个基本的问题。点火延时持续时间将是在引擎正常工作时影响转速,压缩条件和燃油喷射提前角的重要因素,而特殊的燃烧室结构将会是另一个影响因素。CITE-R燃油的最小十六烷值为37,但其已测定的数据只适用在低速状态下,并不能直接应用在6000 rpm 的高转速引擎的预测中。然而,可以从这些有用的数据中,得到小型高速引擎在使用轻油(十六烷值为55)时,以最小可起动压缩比时的表现,估测如表1所示(更高的压缩比会导致额外的热损失和最大气缸压力的增长)。这些数据指出在6000 rpm时使用CITE燃油是可行的。但是在低负荷的情况下,要求入口温度保持在室温下,这对于采用涡轮增压的机型是不可能的。 表 1 达到要求转速时发动机的速度特性* 发动机转速 5000 6000 7000 平均活塞速度, ft/min (m/s) 过 2915 3500 4085 (1480) (1775) (2070)要求的bmep 转速, lb/in2 (kPa) 158 132 113 (1090) (910) (780)预测多缸机的 fmep, lb/in2 (kPa)* 72 83 95 (495) (575) (655)所有发动机的Imep, lb/in2 (kPa) 230 215 208 (1585) (1485) (1435)多缸机的Bsfc , 采用0.33 lb/inp.h(0.20 kg/kW.h), 0.48 0.54 0.61 Lb/inp.h (lb/bhp.h (kg/kW.h) (0.29) (0.33) (0.37)* 2-1/23-1/2 in (88.988.9 mm) 缸径的发动机在 1 bhp/in3 (45.5 kW/cm3) 功率输出。* 见图2燃油喷射时间由燃油喷射系统依靠后置式装置控制,但是在现实中,开发新的喷射系统以替代传统的脉冲式喷油泵系统是不可能的。在一个固定的节流孔喷口,需要考虑的是如何在6000 rpm 下,以全负荷运转时能有60mm3 的喷射量,这就要求能提供合适的特性曲线,调整压缩比在11:1。直喷式的燃烧系统在高转速下延长喷射时间,其结果是很严重,这就不得不采用固定的节流孔喷口。另外一个附加的问题就是大负荷下DI系统工作的困难性,特别是在更高的冒烟界线空燃比情况下,要求更大的功率以达到预定的指标。在Ricardo的早期研究中已经表明,基于慧星漩涡式的燃烧室系统(见表3),工作在4500 rpm 的小型高速民用引擎通过改进可以达到特殊应用的目的。仅管有独立的起动装置和完全多油路系统,但是主要可预见的问题是在高热负荷的情况下,并没有合适的DI系统;尽管如此,随着可应用于多缸机的内置式辅助装置的应用,以及CITE燃油的限制,这些问题都不会很严重。 表 2 发动机进排气系统 - 气门面积和限制速度* 合适的气缸盖布置的形式 气门面积比例 相对于缸径的百分比 结果 进气 排气 Total rpm,max 传统的两气门自然进气 18 12 30 3900 平顶燃烧室四气门涡轮增压 17 17 34 4000 斜屋顶三气门涡轮增压 21 21 42 5000 斜屋顶四气门涡轮增压 25 25 50 6000* 3-1/2 in (88.9 mm) 发动机行程. 受限制的平均进气气体速度: 自然进气,210 ft/s (64 m/s); 涡轮增压, 230 ft/s (70 m/s).在项目开始的时候,一些DI系统被拿来做选择,除此之外有一台4000 rpm的试验用单缸机以及相配套的已设计完成的DI系统版本。在那段时期,受制于噪声,烟气和独特的废气排放法规,分离式燃烧室受到更多的重视,而在DI系统上的试验工作并没有像现在一样被提上议程。引擎的磨擦问题与同样尺寸大小的常规民用多缸高速涡轮增压柴油机相比,很明显,通过提高转速和活塞速度可以显著提高fmep。在表1中可以清楚地看到,如果想到同时达到功率输出和燃油消耗率两项指标是一个很大的问题。图2所示是一台典型的民用引擎fmep/速度曲线图,其估测的fmep是从高速多缸机的试验中获得的。在这个测试中,一些柴油机机构基本的机械磨擦的增值已被假定,直到涡轮增压装置增加气缸压力和采用更大的轴承以获得可接受的可靠性。另外,进排气的泵气损失将会大幅增加高速机的fmep,除非能够设计一个合适的气门直径并长期保持不变。 表3 燃烧室特性 慧星式 直喷式 漩涡式燃烧室 燃烧室特殊转速 (受限于 A/F) 好 失败燃烧控制和机械负荷 好 差热负荷和热损失 差Poor 好独立起动性 差Poor 好采用加热塞的起动性能 好Good -轴针类型 针式 多孔式排放 (NOx) 好Good 差r多燃料性能 差 失败按照汽油标准,该机型的单位机械效率将会很低,但经验告诉我们,尽管在细节方面的设计可以获得额外的收益,但是较低的水平仍然会隐藏在设计的规格里。单缸试验机基于对外形的考虑,单缸试验机的最后设计被确定,其基本尺寸为:缸径冲程: 3-1/2 in 3-1/2 in (88.9mm88.9mm);正常满负荷转速范围,3000-6000 rpm;最高气缸爆发压力,2500 lb/in2(17.3Mpa)。气缸爆发压力也许比传统标准定得稍高,过去的经验显示,设计如此一台引擎是很危险的,所以在试验过程中会造成不可遇见的局限性。事实上,最初可正常工作的DI版本的设计为3000 lb/in2(20.7Mpa),但在随后的慧星版本中有一定的减小。慧星涡旋燃烧室引擎的布局如图3-5所示,整机的主要部件如图6所示,如下所述:曲轴箱曲轴箱及其后盖由球墨铸铁(BS 1452:1961,Grade 14)铸成,两者通过螺栓联接。曲轴箱延用Ricardo设计的E/6 可变压缩比汽油机的曲轴箱,这就导致了燃烧室如E/6型机那样,置于曲轴箱前端。有三个主要轴承,全部采用铅青铜合金衬套,中间一个轴承采用止推轴承。后轴承作为曲轴的延长轴的固定轴承,可以进行调节轴向位置,所以不能分担中间轴承承受的燃烧负荷。曲轴曲轴由一次锻造成形的渗氮钢,BS 970:1955 En 40c。平衡重为整体式,只用来平衡旋转惯性力,同时设有平衡一阶和二阶往复惯性力的平衡轴。所有主轴颈和曲柄销表面渗氮处理,三处主轴颈直径,从前到后,分别为3,3,and 2-3/8 in (76.2,76.2,and 60.4 mm),曲柄销直径为2-5/8 in (66.6mm)。连杆为获得较好的模具和避免多余的费用开支,连杆的方案从民用机型上进行选择,最后选定了福特 2700 系列柴油机的连杆。连杆大头可承受的最大气缸爆发压力为3000 lb/in2,连杆小头则通常不被考虑。因此在将该型连杆运用到慧星漩涡燃烧室版本的引擎上时,要求其最大气缸爆发压力为 2500 lb/in2。轴承采用2700系列引擎上的型号,由15%的锡铝合金做成衬套,而小头衬套采用预制的铝青铜合金。除了在加工时要求仔细抛光和检验连杆质量外,连杆小头应尽可能减小宽度,以减小活塞销座和活塞销衬套的热负荷和惯性负荷。加在连杆大头固定螺栓上的扭矩高于传统的标准,以防止在6000 rpm时,当活塞到达排气上止点时螺栓帽脱落。轴承的提供者The Glacier Metal Co.Ltd. 出了计算机计算的结果,表明所提供的轴承的工作范围是合适的,尽管当高速时连杆大头将承受相当大的惯性力。活塞和活塞销活塞由含硅13%的铝合金整体铸造而成,性能为BS 1490:1970 LM13WP,在斜屋顶燃烧室的有角度的一面上开有浅槽和双凹槽。活塞采用两道气环,第一道为桶形环,第二道为锥形的扭曲环;油环也作相应的设置。为提高耐磨性,在环的磨擦面上镀铜。不镀铬是因为镀铜的环性能已经足够了。虽然活塞高度太大(相对于柴油机的普遍标准)会使设计合适的活塞,活塞环和缸套变得困难,但是在这个机型里活塞仍按照实际需要被特意设计成较高的高度。尽管如此,这一些小问题根据被经验所克服。活塞的冷却和连杆小头的润滑是通过一个安装在曲轴箱内的固定喷嘴定时喷射实现的。这个方法可以取消连杆大头轴瓦处油槽的设计,从而使轴瓦完整。两个活塞的改进,盘式冷却系统的安排,可形变式核心的设计都如图所示。为获得可行的冷却方案,对活塞环带的盘式冷却系统进行设计,承受从活塞头部传过来的气体压力的支杆被设计得近可能细,在这个区域里允许有额外的扭曲。形变式核心有着卓越的性能。活塞销材料为淬火钢,直径为1-3/8 in (34.9mm):缸套和水套因为使用漩涡式燃烧室系统,在局部有较高的热传递速率,伴随有 2500 lb/in2的气缸最大爆发压力,使得在设计湿缸套的时候有相当的难度。根据传统的铁制气缸套厚度设计,则会使第一道环反向点的表面温度额外地上升。最后方案选择使用钢制气缸套,在内壁镀上一层厚度为0.0015 in (38m)的坚硬的铬合金。越向顶部,气缸套越薄,使得温度能得以控制,但是较小的厚度将会削弱缸套的刚度,不利于抵御有水一侧的冲击。缸套的顶端被折边以适当的过盈配合安装在气缸上部,下端装有水封安置在曲轴箱顶部;其径向定位由曲轴箱内的气缸螺栓提供,水封为通常的橡胶环。气缸盖总成气缸盖以及连接架,是发动机最复杂的部件,现在依旧有不少设计问题有待解决。如何设计合适的气门和气道成为其具有相当难度的核心问题。在传统的缸径为5-1/2 in的柴油机中为气门留有3-1/2 in的直径位置,这样可以为漩涡式燃烧室系统的活塞区域提供高达5.7 ihp/in2 (0.66 indicated kW/cm2)的冷却能力。仅管四气门的布置在其它地方得到了充分的验证,但是在这个慧星漩涡式燃烧室做四气门的布置并不十分恰当,因此最后选定了三气门的布置方案。斜屋顶的头部表面积是很重要的,一部分留给必要的气门面积,但更多是用来预留额外的散热空间。通常情况选择使用两个排气门,以减小它们各自的尺寸,同时使用一个进气门。但是最终一个相反的方案被选定,同图8所示。一对进气门布置在中央,因为考虑到如果采用一对排气门的话,较长的排气管会对燃烧室顶部强烈地加热,对排气门设计不利。不对称的燃烧室和气门布置以前也被研究过,但因其不能显示出实际的优势和与多缸机的要求不符合而被舍弃。在这个设计中,通过将加热塞置于燃烧室较低的位置,采用所谓的延展式外形的慧星燃烧室以最小化空间。为了冷却进气门、排气门以及两者之间的桥梁结构,钻孔是不可避免的,这就需要对燃烧室与水套之间的金属层厚度做精确的控制,并且要求水套的表面保持清洁。漩涡燃烧室的加热塞置于窄处,由Nimonic 80A合金精密铸造而成,是一个轻型的定位于铜衬垫的上边缘口,根据经验这样的强化在一些地方需要直接冷却,而不像民用发动机那样存在一个间隙用来迅速暖机。在燃烧室上部有一个由球墨铸铁制成的喷嘴。一个传感器的触点置于气缸顶部的后端。轻型合金机座直接安装在缸套的上边缘,没有用到衬垫,通过8个单头螺柱垂直夹紧在水套的上边缘。根据经验,燃气的冲击不构成问题,设计的单头螺柱满足1.4倍于满负荷燃气冲击的要求。进排气的顶置凸轮轴采用11个螺栓,通过三角架布置在顶部,以确保安全。凸轮轴通过摇臂驱动气门的开启和关闭,气门间隙的调节通过对气门挺柱顶端的垫片来实现。凸轮轴对于进气门关闭和排气打开的时刻的控制是可行的。进气门的材料采用钢BS 970:1955 En 59 “XB”,对于单独的排气门采用特殊的21-4/n奥氏体钢,但根据经验,排气门需要承受高温,在头部会发生穴蚀现象,因此在头部采用Nimonic 80合金,并镶有铬合金。为了实现多缸机在质量上的指标,整机广泛采用铝合金材料,包括气缸盖和机体部分。气缸盖初始设计采用BS 1490:1970 LM25WP Al-Mg-Si合金,但该铸件却被证实在进气口附近存在多孔渗水现象。因此在铸造方面做了相关改进,材料也被改成高温RR 350 Al-Cu-Ni-Co-Sb-Zr 合金。之后铸件就再也没有出现过大的松孔,但随之而来的是进排气门之间的桥梁结构出现破裂现象。冶金实验证明微孔收缩和薄膜都是在铸造中形成的,但是这些缺陷是非常严重的,通过一对固定的热电偶对称置于桥梁结构不同深度处,以选择合适的插入点的方法,因为物理上的原因在这里是不适用的。如表4所示的这些实验中,证实按照最初的设计程序计算得出来的数据,如图4所示,采用Lm25WP 的合金时,材料处于正常工作范围的边缘,采用RR 350时材料失效并不仅仅是热疲劳影响。在对设计做了一定的修改后,加强了桥梁结构的冷却和强度,同时减小了进排气门的直径,分别为1.050 in (26.7mm)和1.505 in (38.2mm),在室温下经过机械负荷测试时,得知缸盖甚至只是在装配后就会发生一定的扭曲变形。微型应变仪被应用在这些桥梁结构,气门座,气缸盖总成,装配,整机的扭矩和气缸体所承受的气缸爆发压力是很合适的。测定的结果,与一台相似的技术成熟的,采用铝制气缸盖的民用两气门柴油机相比较,比较结果见图9,证明在冷态时两种机型的桥梁结构都存在明显的预应力,但是在发动机工作会被热应力负荷所抵消。对RR 350材料的冷态结果分析表明,在温度上升后,有足够的安全系数保留,也没有明显的缺陷问题。 表4 气缸顶部中央的热流和温度* 估测 标准 局部热流, 518,000 475,000 Btu/ft2.h (MW/m2) (1.63) (1.5) 缸壁一侧的温度 504 525 F (C) (262) (279) 推断* 转速: 215 lb/in2 (14.85 bar) at 6000 rpm. 现在设计的气缸盖,已经证明其失效时间为早期设计的2倍,可以说以上所说的问题都已经克服,一个可行的气缸盖设计已经成形。正时驱动平衡齿轮一个后置盖将整个齿轮传动系统罩住,下方是两个主平衡轴传动齿轮和两个二级两倍速的平衡轴传动齿轮,在上方是一个一倍的减速齿轮。平衡轴位于曲柄行程的下方。平衡轴以钢棒材为毛坯,一端磨削后压入薄铁管,使其有光滑的外表面以减少动力损失和机油的搅动。从定时齿轮箱伸出的半速输出轴前端与喷油泵相联,后端与齿形带轮联接。两根顶置凸轮轴通过1:1的同步带轮与位于缸盖总成齿轮的副轴相联接;这根副轴有一定的挠性,以允许将来需要改进时,燃烧室的改变和气门位置的布局能够得以实现。深沟球轴承被应用于正时传动系统。燃油喷射系统在更早的时候,就开始采用传统的高压燃油喷射泵系统,之后又开始对适用在本机型,转速在3000 rpm 的喷射泵的研究。该工作单元是从一台适用于3000 rpm,两冲程三缸机的喷油泵上改进而成的,其单独的动力学原理需要在高速下得以验证。为获得足够高的喷射速度和喷油量,一种新的凸轮轴被开发出来,这样三个单独的工作单元可以相互联系起来。喷油泵通过一个手动的调速器控制,以使得在调节喷油时刻时不用停机。在喷油泵附近靠近水套的地方允许有一短的高压油管,而传统的S号的轴针式喷嘴正好合适。民用发动机上的轴针热屏蔽装置可以用来保持轴针尖较低的温度。冷却回路通过一个外置的电机驱动冷却液高速流动,同时在主循环上还安有一流量计。引擎的冷却回路是一个平行系统,冷却液同时从几个入水口进入。在气缸水套底部开有两个互不干涉的孔,冷却液通过这两个孔经过缸套和缸套边缘的压边处。然后冷却液经12个在缸盖和水套之间的缸盖热圈上的孔进入气缸盖。另一条水路是在气缸顶部的桥梁结构上钻有四个孔,冷却水流在进入主循环前先流经进气门周围的桥梁结构。缸盖通过一个单独的排水通道与排气门相联。润滑油路同样,一个单独的机油泵用来提供润滑油,以满足对润滑和活塞冷却的要求;流量计安置在一条独立的通路上。所有的润滑油经过机体内分离的钻孔和外部管道相联,经验表明在典型的发动机上,这样的布置可以最大程度上减少成本。低压供油给气门结构。机油箱是干式的,同时有一定的加压以防止机油浸到平衡轴上。发动机性能对引擎的燃烧系统和喷射系统的改进并不是同时进行的,而且对发动机的进气系统并没有做最优化设计。因此,这里所引证的数据只是表明指标要求能够达到,而不是为了对发动机性能做最优化设计。当然,更多的成果也是被期待的,特别是在超过转速范围时的平衡性和在超高速下燃烧系统的机械和热负荷数据。测试安装发动机与电子应变测量计相联,也可与驾驶单元相联接。冷却液出口和润滑油入口温度可以通过水冷散热交换器自动控制。压缩空气由独立的空气压缩机提供,中间联有一个ALCOCK滞流空气流量计,另有一个中冷器,用来控制进气温度和减少冲击效应。排气管通过一根短管与一个膨胀室相连,在出口处压力得到控制;在对发动机进排气系统的增压模拟中,一个气室或一排管道被用来进气。其测试安装结果如图10所示。通常的测试方法是,通过改变喷油量,控制引擎工作在一定的负荷范围以上,并保持一定的转速,而进气量和排气管背压保持与对照样机相同。这样并不是模拟增压发动机工作在相同的负荷范围内,而是通过这些数据大致地预测增压发动机的工作状况。根据表现出来的发动机的特征结果,在相同的进气量和背压的情况下,对中断结果和动力磨擦的情况的分析,允许对多缸机指标的推测。由于CITE燃料在英国供给困难,USATAC赞成使用本质上相同的航空用低烷值燃油(D.Eng R.D.2486),其十六烷值为37。一些实验也使用C.I.油料,是一种本质上与ASTM Grade 2-D柴油相似的燃油,其十六烷值为55。测试工作对这台机械部分改进的新型的发动机,大部分测试工作表明其完全达到了215 lb/in2 (1485 kPa) imep的指标要求。一些数据也表明,在达到满负荷工作运转时,转速会降到3000 rpm。在开始测试时,在6000 rpm的转速下其冒烟界线是可以接受的,但在排气管边缘的热电偶测得的排气温度过高,达到了1650F (900C)。这是由于非常长的喷射时间和延迟燃烧造成的。对燃油喷射系统的改进减低了喷嘴的喷射速度,但是只是在名义上有大量的减少,在喷嘴口处,只有20%的减少量在喷射时间内。对喷射系统的简单计算出现另人失望的结果是自然的。因此就需要像一开始一样,对高排气温度时固有的热力问题在高排气温度进行研究和寻找设计解决方案。在负荷范围内,在进气压缩比为1.9和6000 rpm时,两个燃料的性能如图11所示,而其典型的气缸爆发压力和轴针抬起高度如图12所示。从以上数据可得知,增压对于两种燃料达到性能指标都是非常有效的,其在冒烟边界的A/F值为0.055。这些值得注意的数据结果证明,最后燃油注入燃烧室的时刻为40 deg atdc,这可以归功于慧星燃烧系统出色的混合气能力,特别是在活塞凹陷处非常有效的气体流动。在使用轻油时,慧星燃烧室可以达到最适宜的起动性能,但在使用对点火要求更低要求的航空低辛烷值燃油时,它就不可能达到在相同的进气压力下相同的燃烧效果,不可避免地会有些损失。在小负荷时不能发火的问题同样也在使用航空低辛烷值燃油时体现出来,增加一定的进气温度可能成为必要的措施多缸机使用一根进气总管采用对冷却液特殊控制的二次冷却器成为必要的手段。从气缸爆发压力图表上可得知,气缸最大爆发压力在高速的情况下,可以测得其值在满负荷时为1475 lb/in2 (10.2MPa)。在恒定的进气压力和最适宜的喷油时刻时,使用轻油的速度曲线如图13所示,喷油时刻的改变如图11所示。应该指出在燃烧室低转速输出时相对于高转速下的要求有所降低,尽管随着更大的改进也会有相应的改进。现在相当小的气门被应用,而早期的进气门关闭时刻为40 deg abdc,其体积效率在5000 rpm时下降,但是多缸机对扭矩的储备要求还是可以接受的。多缸机的可能性在对单缸机试验完成时,还没有认为值得花时间对多缸机的进行设计研究。但是,为了获得对今后整机样式的大致了解,一些可能的四缸机方案草图还是有的,其前端视图和侧视图如图14所视。该发动机的正视图表现了单列、干机油箱的设计,它的气缸盖除了采用整体式缸盖外,本质上与单缸机的相同。涡轮增压器安装在飞轮上方,与进入总管和中冷器相联接。该引擎,包括各个辅配件,包括飞轮长31-3/4 in (806 mm),高 32-3/4 in (831) mm),宽24 in (609 mm),单位质量功率为9.3 bhp/ft3 (245 kW/m3)。为了达到特定的重量要求,其干重量必需达到470 lb (213 kg)。其质量与排量比为3.49 lb/in3 (0.096 kg/cm3),而传统的车用柴油机的质量与排量比为3-5 lb/in3 (0.08-0.14 kg/cm3)。基于单缸试验机的实验数据,为该机型设计的制动器的性能也基本上达标了。在4000 rpm时要求有10%的扭矩储备要求,因此,作为一台涡轮增压发动机,已经达到了扭矩的要求。在4000 rpm以下,供油量需要削减以保持其排烟有约5%的不透明度,这样在低速时就可能获得一定的性能增长。进气压缩比被要求在通过中冷器损失后还有2.7,其最大的气缸爆发压力为1950 lb/in2 (13.5 MPa)。排放情况在单缸机上并没有做测定,但是慧星漩涡式燃烧室已经被证明其拥用极低的排放,特别是是氮氧化物方面,倘若不发火的问题能够得到解决,那么这台发动机满足1975C.A.R.B的要求也就成为可能。概括和总结 在USATAC发起的设计和测试过程中,达到高效的输出1 bhp/in3 (45.5 kW/cm3)和低重量的柴油机的可能性也曾被研究过,采用高转速和一般的活塞运动速度,分别为6000 rpm和3500 ft/min (17.75m/s)。预测要达到的目标是imeps达到 215 lb/in2 (1485 kPa),因此也就需要采用涡轮增压系统。 在研究一开始所涉及到的三个主要问题:1. 引擎的进排气系统,主要是尽可能提供比传统发动机更大的气门面积。2. 燃烧,包括最基本的在很短的时间间隔内实现压燃的问题,也有燃油喷射系统和燃烧室的选型。3. 发动机的磨擦,对bfc的控制。从以上几点考虑,一个单缸试验机的设计提了出来,包括对进排气系统和燃烧等细节问题的研究。1. 使用Ricardo设计的慧星漩涡燃烧室系统。2. 采用斜屋顶燃烧室顶部设计,以便提供足够的气门安装空间。3. 双进气门和单排气门设计。4. 合理地对机械部分做出设计,以应付在特殊地试验条件下对很高气体爆发压力的抵抗要求,以及设计适应高热应力的冷却系统。 测试工作表明设计的指标可以达到,尽管改进工作还没有完成,但是试验数据显示进排气系统和采用慧星燃烧室的燃烧系统可以实现转速6000 rpm,这还是在采用传统的高压燃油喷射泵的情况下。采用CITE燃油也可能是出现这种情况的原因。 最初为军事目的设计的,背包大小,特殊功率要求的4缸试验机已经制造完成。致谢 作者非常感谢文中所提及的美军坦克研发中心在这个项目的大力支持,也很感谢他们和Ricardo & Co.对论文出版的许可。参考文献1. B.W.Millington and E.R.Hartles, “Frictional Losses inDiesel Engine.” SAE Transactions, Vol.77 (1968), paper 680590.2. C.J.Walder, “Some Problems Encountered in the Design and Development of High Speed Diesel Engines.” SAE Transactions, Vol.74(1966), paper 650025.3. W.T.Lyn and E.Valdmanis, “The Effects of Physical Factors on Ignition Delay.” Paper 689192 presented at SAE Automotive Engineering Congress, Detroit, January 1968. 4. C.C.J.French, “Taking the heat off the Highly Boosted Diesel.” SAE Transactions, Vol.78 (1969), paper 690463.5. C.J.Walder, “The Reduction of Emissions from Diesel Engines.” Paper 7320214 presented at SAE Automotive Engineering Congress, Detroit, January 1973.21 Toward Higher Speeds and Outputs From the Small Diesel Engine D.BroomeRicardo & Co.Engineers(1927) Ltd.(England)THE AUTHORS company has long been concerned with the development of the small, high-speed diesel engine, and is particularly associated with combustion systems for this type of engine. Although such engines are not common in the North American continent, production and use in Europe and Japan is considerable, totaling several million units. These are, typically, naturally aspirated 4-cyl engines of 25-35 in3 (400-600cm3) displacement per cylinder, operating up to speeds of 4000-5000 rpm, with a limiting piston speed of about 2400ft/min(12m/s). In discussion with the U.S.Army Tank-Automotive Command (USATAC) at , Mich.it was proposed that the military requirement of high power from a small lightweight package could be achieved by exploiting higher speeds than hitherto, rather than the application of increased levels of turbocharger alone, and this led to the formulation of a research program to study combustion and breathing problems under such conditions. This paper describes the work carried out to date, which has involved the design, manufacture, and preliminary test work on special single cylinder engine. THE PROJECT The project specifications finally laid down by USATAC can be summarized as follows:1. Design, procure, build, and test a single-cylinder engine of 3-1/2 in (88.9mm) bore and stroke, to operate at the highest possible speed, but certainly above 5000 rpm. Simulation of turbocharged conditions to be achieved using a separate air supply.2. To develop the single-cylinder test engine to achieve performance targets such that a 4-cyl version for military duties could produce 1 bhp/in3 (45.5kW/cm3) displacement with a target dry weight of about 3.5 lb/bhp (2.13kg/kW).3. The design not to be influenced by conventional practices, with the aim of minimizing mechanical and thermal stresses.4. Operation on CITE-R fuel (MIL-F-46005A (MR) to be the primary requirement. Initially, fuels down to aviation gasoline were to be investigated, but this latter requirement was subsequently relaxed.5. Lubricating oils to the MIL-L-2104B specification to be used if at all possible.6. The final phase of the project to include a design study for a 4-cyl military engine, embodying the lessons learned on the single-cylinder test unit.7. Starting, idling, and light-load operation of the multicylinder engine must not be compromised.PRELIMINARY DESIGN CONSIDERATIONS A simple examination of the cylinder size and power output target rapidly showed the limitations that the maximum engine speed would have on performance (Table 1). Starting from the minimum speed specified of 5000 rpm, it is clear that speeds of 6000 rpm and above entail piston speeds equal to those of racing gasoline engines. While the reductions in bmep through use of high speeds are significant, the increases in fmep (estimated from past results obtained at the authors company, much of which has been summarized in Ref.1) give very little return in reduced imep. Naturally aspirated automotive diesel engines working to the strict smoke limits of a few years hence can only operate up to about 145lb/in2 (1000kPa) imep; hence it was clear that some measure of turbocharging would be required. A further penalty of high speed and high engine friction is in fuel consumption, and Table 1 makes clear how the bsfc would worsen rapidly to levels no better than a gasoline engine, so losing one of the major advantages of the compression ignition cycle. In these circumstances,it was decided to limit the speed of the research engine to 6000 rpm. The major performance problems involved in the design of an engine to meet these requirements might be summarized as follows: ENGINE BREATHING-Previous experience on small high-speed diesels had shown that the major limitation on imep at high piston speeds is the breathing of the engine (2).Hence, valves of sufficient flow area had to be provided to allow efficient operation up to 3500 ft/min (17.8m/s) piston speed, some 50% higher than levels normally employed in diesel engines. This would certainly require departures from conventional cylinder head arrangements, involving inclined multiple-valve designs (Table 2);turbocharged operation brings a slight bonus in that the higher inlet air temperatures minimize pressure losses and reduce volumetric efficiency changes.In addition, possible turbocharger matching requirements had to be borne in mind. While, for automotive engines, torque backup requirements normally favor minimizing the available boost at the rated speed, so that a large exhaust valve area is not mandatory, in this case the very short absolute exhaust gas release periods suggested that the exhaust mean gas velocities should be kept low, and exhaust valve area about equal to that of the inlet. Also requiring consideration was the question of valve timings. For the inlet, high speeds are normally associated with a late closing point, yet in the case of a diesel, and with closing points later than about 45 deg abdc, there would be a progressive sacrifice in starting ability, as well as some loss of low-speed performance, which would further impair the natural torque backup characteristics of the engine. For the exhaust, the turbocharger matching requirement again dictates an early release of the gases on the expansion stroke, and timings later than about 60 deg abdc do not show to advantage at high speeds. While a long overlap period could contribute to reduction of exhaust gas and exhaust system component temperatures, such gains would be minimal at high speeds due to the very low quantity of scavenge air which might be passed relative to the trapped flow, and the mechanical problems of obtaining the piston/valve clearance would place a severe penalty on the combustion system. COMBUSTION PROBLEMS-A fundamental problem likely to affect the engine at the speeds contemplated was the likely duration of the ignition delay period. Ignition delay is a function of engine speed, compression conditions, and injection timing for a fuel of particular ignition delay is a function of engine speed, compression conditions, and injection timing for a fuel of particular ignition quality at normal running conditions (3),if factors related to the particular combustion chamber configuration in use are considered as being of second order. CITE-R fuel has a minimum specified cetane rating of 37,but the published data on engine delay using this fuel covered only low-speed conditions, and were not of direct use in predicting results at 6000 rpm. However, consideration of these available data, together with the known performance of small high-speed engines operating up to 5000 rpm on gas oil (55 cetane), led to the estimates shown in Fig.1,for the lowest compression ratio which would allow acceptable starting (higher ratios would give excessive heat losses and maximum cylinder pressures).These suggested that unaided or true compression ignition operation at 6000 rpm was feasible on CITE fuel, although the light-load condition would require inlet manifold air temperatures to be maintained significantly above ambient-not an impossible requirement for a turbocharged engine. Injection periods being controlled by the injection system will depend on the latters type, but for practical reasons there could be no possibility of developing new systems for the project, to replace the conventional jerk pump arrangement. With a fixed orifice area nozzle, there would be considerable problems in passing the required full load quantity of up to about 60 mm3/injection at 6000 rpm at the required rate, yet obtaining satisfactory characteristics for idling, the turndown ratio being about 11:1.The effect of an extended injection period on combustion at the high speeds required could be very severe on a direct injection (DI) combustion system, where use of a fixed orifice nozzle would be inevitable. In addition to this problem the major difficulty of the DI was seen as the high mechanical loading, accentuated by the higher smoke-limited fuel/air ratios (A/F) requiring higher boosts to achieve the target rating. While Ricardos earlier research work had shown that DI systems could be made to operate up to 4500 rpm naturally aspirated, on balance (see Table 3) the Comet swirl chamber system, developed over many years for the small high-speed commercial engine, was considered to offer greater potential for this particular application. The major problem foreseen was high thermal loading, although the unaided starting and the full multifuel capabilities were also less satisfactory than those of the DI: however, with built-in aids such as could be applied to the multicylinder engine, and with a restriction to CITE fuel, these latter were not considered to be too serious.At the start of the project, then, some consideration was given to the DI as an alternative, and in addition to test running under boosted conditions of an existing 4000 rpm single-cylinder research unit, designs were completed for a DI version of the test engine. Since that date, the increased pressure of noise, smoke, and particularly exhaust emissions legislation, has increasingly favored the divided chamber system, and test work on the DI version is not now likely to take place.ENGINE FRICTION-Comparing the proposed multicylinder high-speed turbocharged engine with a conventional commercial engine of the same cylinder size and number, it was clear that the former would have a significantly higher fmep through the use of higher rotational and mean piston speeds. As already made clear in Table 1,this would pose serious problems in relation to the attainment of both the target output and an acceptable fuel consumption.Fig.2 shows how using a typical commercial engine fmep/speed curve from Ref.1, the estimated fmep of the high-speed multicylinder engine was obtained. In this estimate, some increase in the mechanical friction of the basic engine structure was assumed, since the turbocharged condition would increase cylinder pressures and require larger bearings to give acceptable reliability. In addition, inlet and exhaust pumping losses could add materially to the high-speed fmep, unless acceptable valve sizes could be maintained. By gasoline standards,then,the mechanical efficiency of the unit would be poor,but experience had shown that although attention to detail throughout the design could yield gains,these low levels were implicit in the project specification.SINGLE-CYLINDER TEST ENGINE Based on the considerations outlined, the definitive single-cylinder test engine was designed, the boundary operating conditions for the engine being: bore and stroke,3-1/2 in 3-1/2 in (88.9mm88.9mm);normal full-load speed range,3000-6000 rpm; and maximum cylinder pressure,2500 lb/in2(17.3Mpa). While the cylinder pressure limit may seem high by conventional standards, past experience has shown the dangers of designing such engines to low limits, and thus inflicting unforeseen limitations on the test program. In fact, originally, with possible work on a DI version in mind, a limit of 3000 lb/in2(20.7Mpa) was set, but as noted, this limit was later reduced for the Comet version. The Comet swirl chamber engine layout is illustrated in Figs.3-5, and the complete engine shown in Fig.6.Of the major components, the following may be said: CRANKCASE-The crankcase and rear-mounted timing case and cover are in gray flake graphite iron to BS 1452:1961 Grade 14,spigoted or doweled and bolted together. The crankcase design was adopted from that of the Ricardo E/6 variable compression ratio gasoline engine, which results in the presence of the front chamber of the crankcase unit, where the E/6 timing drive was situated. Three main bearings are use, all of lead-bronze bushing type, the center bearing being the thrust bearing. The rear bearing acts only as a steady bearing to the otherwise long extension of the crankshaft, the clearance being adjusted so that it cannot take the firing load off the center bearing. CRANKSHAFT-The crankshaft is a one-piece forging in nitriding steel to BS 970:1955 En 40c.The balance weights are integral, and balance only the rotating loads, since primary and secondary balancer shafts are fitted to the engine. All journal and pin surfaces are nitrided, the diameters of the three journals being 3,3,and 2-3/8 in (76.2,76.2,and 60.4 mm),respectively, from front to rear, and of the pin 2-5/8 in (66.6mm). CONNECTING ROD-To obtain the better material properties associated with a forging without the expense of special dies, a search was made of commercial engines, and the connecting rod of the Ford 2700 series diesel engine finally selected as the most appropriate. While satisfactory big-end bearing loadings were achieved at the 3000 lb/in2 maximum cylinder pressures, the little-end design was considered inadequate, and the decision made to use this rod only for the Comet swirl chamber version of the engine, at a pressure limit of 2500 lb/in2. The bearings are as used on the 2700 series engine, that is ,15% reticular tin/aluminium half liners, the little-end bushing being a wrapped lead-bronze item. In addition to careful checking and polishing of the rod, the little end is reduced in width, the better to distribute the firing and inertia loads between the piston pin bosses and the little-end bushing. A higher torque than standard is used on the big-end setscrews, to prevent the cap lifting off due to the inertia forces at tdc exhaust, at 6000 rpm. Computer calculations carried out by the bearing suppliers, The Glacier Metal Co.Ltd., showed that the proposed bearing arrangements were acceptable, although the big end in particular has to accept very arduous conditions at high speeds due to the great inertia of the relatively massive connecting rod (Fig.7).The importance of correct form for both the pin and the bearing under these conditions cannot be overstressed; this apart, the only problem that occurred was rapid cavitation attack in the top (loaded) half liner with the original clearance. The cause of this is evident in Fig.7,and a reduction in clearance to 0.0022 in (56m) cured this trouble. PISTON AND WRIST PIN-The piston is a one-piece sand casting in 13% silicon aluminum alloy to BS 1490:1970 LM13WP,with the shallow trench and twin recesses of the Comet combustion system formed in one face of the angled (pent-roof) crown surface. Two compression rings are used, the top being a plain barrel-faced ring and the second a taper-faced internally stepped (twisted) ring; the slotted oil-control ring is of the conformable type. Rings are supplied copper-plated on the rubbing faces to assist in bedding in, but are not chrome-plated, since this facing is applied to the liner. The piston was designed deliberately of relatively great height, since it was feared that the very high (by diesel standards) piston speeds together with boosted operation would create difficulties in obtaining acceptable piston, ring, and liner conditions, and it was not thought desirable to accentuate problems more than was necessary. However, relatively little trouble has been experienced with the ring pack. Piston cooling and little-end lubrication is via an oil spray from a fixed jet located in the crankcase. This method was selected to avoid grooving the big-end bearing liner, which would have reduced its capacity. Two piston designs were developed, a tray-cooled arrangement and the soluble core design shown in the figures. To obtain acceptable cooling of the ring belt with the tray-cooled design, the struts transmitting gas loads from the crown to the wrist pin bosses were thinned as far as was thought practicable, but this arrangement was found to allow excessive distortion. The soluble core design has given excellent service to date. The wrist pin is of case-hardened steel,1-3/8 in (34.9mm) in diameter: CYLINDER LINER AND WATER JACKET-The high rates of local heat transfer associated with the use of a swirl chamber combustion system, together with the 2500 lb/in2 maximum cylinder pressure limit, led to design difficulties with the wet-type cylinder liner, since calculations showed that a conventional iron liner of thickness adequate to withstand the gas loads would give excessive surface temperatures for acceptable lubrication at the top ring reversal point. The solution adopted was to use a steel liner, with the bore given the necessary surface finish before being plated with hard chrome to a thickness of 0.0015 in (38m) by the Chromard process. Toward the top, the liner is thinned to provide the necessary temperature control, while the greater thickness lower down enhances rigidity to combat water-side attack. The liner is flanged at the top and seats on the cylindrical mild steel water jacket itself seating on top of the crankcase; radial location is provided by the liner spigoting in the crankcase, a water seal being obtained in the normal way by rubber O-rings. CYLINDER HEAD ASSEMBLY-The cylinder head, with its associated cambox, is the most complex single assembly of the engine, and presented considerable design problems. The most pressing of these centered on the provision of adequate valves and ports-the difficulties here may be appreciated when it is realized that ports suitable for a conventional engine of 5-1/2 in bore had to be provided on a 3-1/2 in bore-together with adequate cooling for the very high rating of 5.7 ihp/in2 (0.66 indicated kW/cm2) of piston area, this with a swirl chamber comber combustion system. The position of the Comet swirl chamber at the edge of the bore does not render the use of four valves very attractive, and although this and other possible layouts were examined, a 3-valve arrangement was finally adopted. A pent-roof head surface was necessary (Table 2), partly to obtain the necessary valve area but primarily to prevent excessive congestion higher up in the head. It is normally preferable to pair the exhaust valves to reduce their individual size and use a single large valve, but the opposite layout was finally chosen, as shown in Fig.8, since the paired valves had to lie in the center of the head and the intense heating of the head from the long port duct if this latter were the exhaust was considered unacceptable. Asymmetrical chamber/valye layouts were also investigated but rejected as offering no real advantages and being incompatible with multicylinder requirements. The so-called externally inserted form of the Comet chamber was adopted to minimize the space occupied by the hot plug forming the lower portion of the chamber. To cool the resulting four bridges ins the lower deck of the head, between the chamber and the inlet valves, and the inlet and exhaust valves, drillings were provided, giving an accurately controllable metal thickness between the hot gases and the coolant, and a clean surface on the coolant side. The swirl chamber hot plug carrying the throat, and made from Nimonic 80A alloy by precision casting, is a light fit on its sides as well as locating on a copper gasket on its upper flange, since experience showed that at these high ratings some direct cooling was necessary, unlike commercial engines where an air gap is used to improve warm up after starting. The upper part of the chamber carrying the injector is a spheroidal graphite iron casting. A transducer tapping into the cylinder is provided at the front end of the head. The light alloy head seats directly on the steel liner flange, no gasket being employed, and is clamped by eight suds rooted high in the head and passing vertically downward through the water jacket top flange. No difficulties with gas blow have been exper
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