專用鉆床液壓系統(tǒng)設(shè)計(jì)
專用鉆床液壓系統(tǒng)設(shè)計(jì),專用鉆床液壓系統(tǒng)設(shè)計(jì),專用,鉆床,液壓,系統(tǒng),設(shè)計(jì)
東華理工大學(xué)長(zhǎng)江學(xué)院畢業(yè)設(shè)計(jì)（論文） 扉頁(yè) XX大學(xué)課程設(shè)計(jì)（論文）課 題 名 稱： 專用鉆床液壓系統(tǒng)設(shè)計(jì) 專 業(yè) 班 級(jí)： 學(xué) 生 姓 名： 指 導(dǎo) 教 師： 30 摘 要作為現(xiàn)代機(jī)械設(shè)備實(shí)現(xiàn)傳動(dòng)與控制的重要技術(shù)手段，液壓技術(shù)在國(guó)民經(jīng)濟(jì)各領(lǐng)域得到了廣泛的應(yīng)用。與其他傳動(dòng)控制技術(shù)相比，液壓技術(shù)具有能量密度高配置靈活方便調(diào)速范圍大工作平穩(wěn)且快速性好易于控制并過(guò)載保護(hù)易于實(shí)現(xiàn)自動(dòng)化和機(jī)電液一體化整合系統(tǒng)設(shè)計(jì)制造和使用維護(hù)方便等多種顯著的技術(shù)優(yōu)勢(shì)，因而使其成為現(xiàn)代機(jī)械工程的基本技術(shù)構(gòu)成和現(xiàn)代控制工程的基本技術(shù)要素。本課題研究的主要內(nèi)容是專用鉆床液壓系統(tǒng)的設(shè)計(jì)及PLC設(shè)計(jì)。液壓系統(tǒng)的設(shè)計(jì)是整個(gè)機(jī)器設(shè)計(jì)的一部分，它的任務(wù)是根據(jù)機(jī)器的用途、特點(diǎn)和要求，利用液壓傳動(dòng)的基本原理，擬定出合理的液壓系統(tǒng)圖，再經(jīng)過(guò)必要的計(jì)算來(lái)確定液壓系統(tǒng)的參數(shù)，然后按照這些參數(shù)來(lái)選用液壓元件的規(guī)格和進(jìn)行系統(tǒng)的結(jié)構(gòu)設(shè)計(jì)，最后對(duì)液壓系統(tǒng)的主要性能進(jìn)行驗(yàn)算。關(guān)鍵字： 專用鉆床；液壓系統(tǒng)；PLC設(shè)計(jì) ABSTRACTAs one of the modern mechanical equipment to achieve an important technical means of power transmission and control, hydraulic technology has been widely applied in various fields of the national economy. Compared with other drive control technology, hydraulic technology with high energy density, convenient and flexible configuration, wide range of speed regulation, smooth and fast is good, easy to control and overload protection, easy to realize automation and electromechanical integration, system design and manufacture and use and maintenance is convenient wait for a variety of significant technical advantages, which make it become the basic constitution the technology of modern mechanical engineering and modern control engineering of the basic elements of technology.The main content of this paper is the design and PLC design of special drilling machine hydraulic system. The design of the hydraulic system is a part of the whole design of the machine, it is the task of the application of the machine, according to the characteristics and requirements, using the basic principles of hydraulic transmission, and draw up a reasonable hydraulic system diagram, and then through the necessary calculations to determine the parameters of hydraulic system, structure design and then according to these parameters to the selection of hydraulic components of the specification and the last, checking the main performance of the hydraulic system.Keywords: special drill press; hydraulic system; PLC design 目 錄摘 要IABSTRACTII1 緒 論12 設(shè)計(jì)任務(wù)書(shū)23 液壓系統(tǒng)的功能原理計(jì)算33.1液壓缸液壓系統(tǒng)設(shè)計(jì)要求分析33.2 負(fù)載分析43.2.1 工作負(fù)載43.2.2 摩擦負(fù)載43.2.3 慣性負(fù)載43.2.4 液壓缸在各階段的負(fù)載值53.2.5 負(fù)載圖與速度圖的繪制53.3 液壓缸主要參數(shù)的確定63.4 計(jì)算和確定液壓缸的主要尺寸63.5 制定液壓回路方案，擬定液壓系統(tǒng)原理圖173.6液壓系統(tǒng)的工作原理183.7 計(jì)算與選擇液壓元件193.7.1 液壓泵及驅(qū)動(dòng)電機(jī)計(jì)算與選定193.7.2 液壓控制閥和液壓輔助元件的選定203.7.3油管的選擇213.7.4液壓系統(tǒng)的驗(yàn)算234 設(shè)計(jì)小結(jié)28致 謝29參考文獻(xiàn)30 1 緒 論本課題來(lái)源于生產(chǎn)實(shí)踐，液壓系統(tǒng)利用液壓泵將原動(dòng)機(jī)的機(jī)械能轉(zhuǎn)換為液體的壓力能，通過(guò)液體壓力能的變化來(lái)傳遞能量，經(jīng)過(guò)各種控制閥和管路的傳遞，借助于液壓執(zhí)行元件(液壓缸或馬達(dá))把液體壓力能轉(zhuǎn)換為機(jī)械能，從而驅(qū)動(dòng)工作機(jī)構(gòu)，實(shí)現(xiàn)直線往復(fù)運(yùn)動(dòng)和回轉(zhuǎn)運(yùn)動(dòng)。其中的液體稱為工作介質(zhì)，一般為礦物油，它的作用和機(jī)械傳動(dòng)中的皮帶、傳動(dòng)它是以液壓油為工作介質(zhì)，通過(guò)動(dòng)力元件將原動(dòng)機(jī)的機(jī)械能變?yōu)橐簤河偷膲毫︽湕l和齒輪等傳動(dòng)元件相類似。液壓能，再通過(guò)控制元件，然后借助執(zhí)行元件將壓力能轉(zhuǎn)換為機(jī)械能，驅(qū)動(dòng)負(fù)載實(shí)現(xiàn)直線或回轉(zhuǎn)運(yùn)動(dòng)，且通過(guò)對(duì)控制元件擾動(dòng)時(shí)，執(zhí)行元件的輸出量一般要偏離原有調(diào)定值，產(chǎn)生一定的誤差。液壓系統(tǒng)主要由：動(dòng)力元件（油泵）、執(zhí)行元件（油缸或液壓馬達(dá)）、控制元件（各種閥）、輔助元件和工作介質(zhì)等五部分組成。在液壓傳動(dòng)中，液壓油缸就是一個(gè)最簡(jiǎn)單而又比較完整的液壓傳動(dòng)系統(tǒng)，分析它的工作過(guò)程，可以清楚的了解液壓傳動(dòng)的基本原理。液壓傳動(dòng)有許多突出的優(yōu)點(diǎn)，因此它的應(yīng)用非常廣泛。未來(lái)社會(huì)是一個(gè)環(huán)保的，低污染，低消耗的社會(huì)，這就要求我們?cè)诟纳埔簤合到y(tǒng)的技術(shù)方面下功夫，作為即將走進(jìn)社會(huì)的我們更應(yīng)該關(guān)注新技術(shù)的應(yīng)用和開(kāi)發(fā)。 2 設(shè)計(jì)任務(wù)書(shū)設(shè)計(jì)一專用鉆床的液壓系統(tǒng)，本題目為新課題，培養(yǎng)學(xué)生綜合應(yīng)用所學(xué)知識(shí)，結(jié)合實(shí)踐知識(shí)，初步具有設(shè)計(jì)一個(gè)中等復(fù)雜液壓系統(tǒng)的能力。學(xué)會(huì)擬訂液壓系統(tǒng)設(shè)計(jì)方案，完成油路塊結(jié)構(gòu)設(shè)計(jì)，提高設(shè)計(jì)能力。培養(yǎng)學(xué)生查閱有關(guān)機(jī)械設(shè)計(jì)手冊(cè)、資料的能力。進(jìn)一步培養(yǎng)學(xué)生制圖，編寫(xiě)技術(shù)文件等基本技能。 3 液壓系統(tǒng)的功能原理計(jì)算液壓系統(tǒng)設(shè)計(jì)是指組成一個(gè)新的能量傳遞系統(tǒng)，以完成一項(xiàng)專門(mén)的任務(wù)。系統(tǒng)功能原理設(shè)計(jì)是根據(jù)主機(jī)的工藝目的或用途、工作循環(huán)、負(fù)載條件和主要技術(shù)要求，通過(guò)配置執(zhí)行元件，負(fù)載分析、運(yùn)動(dòng)分析及編制執(zhí)行元件的工況圖，對(duì)同類主機(jī)及其傳動(dòng)系統(tǒng)的分析比較，選擇設(shè)計(jì)參數(shù)，確定液壓系統(tǒng)的工作壓力、流量和執(zhí)行元件主要幾何參數(shù)等，擬定液壓系統(tǒng)方案和傳動(dòng)系統(tǒng)原理圖，并對(duì)組成系統(tǒng)的各標(biāo)準(zhǔn)液壓元件輔件進(jìn)行選型，最后對(duì)液壓系統(tǒng)的主要性能（壓力損失、發(fā)熱溫升等）進(jìn)行驗(yàn)算。 3.1液壓缸液壓系統(tǒng)設(shè)計(jì)要求分析 設(shè)計(jì)題目設(shè)計(jì)工作循環(huán)為：快進(jìn)工進(jìn)死擋鐵停留快退停止。1.已知參數(shù) 設(shè)計(jì)一專用鉆床的液壓系統(tǒng)，設(shè)計(jì)工作循環(huán)為：快進(jìn)工進(jìn)死擋鐵停留快退停止。主要性能參數(shù)與性能要求如下：切削力FL=14000N；運(yùn)動(dòng)部件自重G=6000N；快進(jìn)、快退速度5.5m/min，工進(jìn)速度60-1000m m/min；快進(jìn)行程L1=250mm，工進(jìn)行程L2=100mm；加速和減速時(shí)間t=0.2s。2 明確設(shè)計(jì)要求 該液壓系統(tǒng)的功率較大，空行程和加壓行程速度差異較大，因此要求功率利用合理。且該系統(tǒng)的壓制力較大，因此對(duì)于工作的平穩(wěn)性、安全性要求較大。 3 設(shè)計(jì)方案 根據(jù)已知參數(shù)和表2-1所示液壓系統(tǒng)工作臺(tái)的執(zhí)行元件為單桿活塞缸，工件的夾緊為液壓缸。 活塞桿繪制工作循環(huán)圖3.2 負(fù)載分析 3.2.1 工作負(fù)載 工作負(fù)載Fe 液壓缸的常見(jiàn)工作負(fù)載有重力、切削力、擠壓力等。阻力負(fù)載為正，超越負(fù)載為負(fù)。 自重G=6000N切削力F=FL=14000N 3.2.2 摩擦負(fù)載 假設(shè)靜摩擦系數(shù)fs=0.2，動(dòng)摩擦系數(shù)fd=0.13.2.3 慣性負(fù)載 慣性負(fù)載Fi 慣性負(fù)載時(shí)運(yùn)動(dòng)部件在啟動(dòng)和制動(dòng)過(guò)程中的慣性力，其平均值可按下式計(jì)算 Fi =G/g*v/t (N) 式中 g=重力加速度， m/s2，g=9.8m/s2 v=速度變化量， m/s2 t=啟動(dòng)或制動(dòng)時(shí)間，s 一般機(jī)械t =0.10.5s，取u=0.1m/s ，t=0.2 s3.2.4 液壓缸在各階段的負(fù)載值（1） 查液壓缸的機(jī)械效率，可計(jì)算出液壓缸在各工作階段的負(fù)載情況，如下表表1所示：表1 液壓缸各階段的負(fù)載情況工況負(fù)載計(jì)算公式液壓缸負(fù)載液壓缸推力/N啟動(dòng)12001333.33加速66007333.33快進(jìn)600666.67工進(jìn)1460016222.22快退600666.6673.2.5 負(fù)載圖與速度圖的繪制根據(jù)工況負(fù)載和以知速度和及行程S，可繪制負(fù)載圖和速度圖，如下圖（圖1、圖2）所示：圖1(負(fù)載圖)圖2（速度圖）3.3 液壓缸主要參數(shù)的確定（1）液壓缸的內(nèi)徑和活塞桿的內(nèi)徑表3-1 按負(fù)載選擇工作壓力1負(fù)載/ KN50工作壓力/MPa 0.811.522.5334455表3-2 各種機(jī)械常用的系統(tǒng)工作壓力1機(jī)械類型機(jī) 床農(nóng)業(yè)機(jī)械工程機(jī)械建筑機(jī)械液壓鑿巖機(jī)專用鉆床大中型挖掘機(jī)重型機(jī)械起重運(yùn)輸機(jī)械磨床組合機(jī)床龍門(mén)刨床拉床工作壓力/MPa0.82352881010182032初選系統(tǒng)壓力P=16Mpa3.4 計(jì)算和確定液壓缸的主要尺寸1 液壓缸缸徑的計(jì)算內(nèi)徑D可按下列公式初步計(jì)算：液壓缸的負(fù)載為推力 式（3-1）式中 液壓缸實(shí)際使用推力14600（N）；液壓缸的總效率，一般取=0709；計(jì)算=0.9；液壓缸的供油壓力，一般為系統(tǒng)壓力（MPa）本次設(shè)計(jì)中液壓缸已知系統(tǒng)壓力=16MPa；考慮到專用鉆床工作時(shí)可能會(huì)超載，故根據(jù)實(shí)際需要，稍微取查缸筒內(nèi)徑系列/mm(GB/T 2348-1993)可以取為40mm。液壓氣動(dòng)系統(tǒng)及元件 缸內(nèi)徑及活塞桿外徑 標(biāo)準(zhǔn)編號(hào)：GB/T 2348-1993表 GB/T 2348-1993 直徑系列直徑系列/mm(GB/T 2348-1993)4、5、6、8、10、12、16、18、20、22、25、28、32、36、40、45、50、56、63、70、80、90、100、110、125、140、160、180、200、220、250、280、320、360由于快進(jìn)速度和快退速度相等，屬于差動(dòng)連接，可以得到d=0.707D,代入計(jì)算并取標(biāo)準(zhǔn)直得d=28.107mm，根據(jù)標(biāo)準(zhǔn)系列，取d=28mm2活塞寬度的確定活塞的寬度一般取=（0.6-1.0）即=（0.6-1.0）40=（24-40）mm取=32mm3缸體長(zhǎng)度的確定液壓缸缸體內(nèi)部的長(zhǎng)度應(yīng)等于活塞的行程與活塞寬度的和。缸體外部尺寸還要考慮到兩端端蓋的厚度，一般液壓缸缸體的長(zhǎng)度不應(yīng)大于缸體內(nèi)徑的20-30倍。即：缸體內(nèi)部長(zhǎng)度快進(jìn)行程L1=250mm，工進(jìn)行程L2=100mm 油缸的有效行程為L(zhǎng)=L1+L2=250mm+100mm=350mm4缸筒壁厚的計(jì)算在中、低壓系統(tǒng)中，液壓缸的壁厚基本上由結(jié)構(gòu)和工藝上的要求確定，壁厚通常都能滿足強(qiáng)度要求，一般不需要計(jì)算。但是，當(dāng)液壓缸的工作壓力較高和缸筒內(nèi)徑較大時(shí)，必須進(jìn)行強(qiáng)度校核。當(dāng)時(shí)，稱為薄壁缸筒，按材料力學(xué)薄壁圓筒公式計(jì)算，計(jì)算公式為 式（3-2） 式中，缸筒內(nèi)最高壓力； 缸筒材料的許用壓力。=, 為材料的抗拉強(qiáng)度，n為安全系數(shù)，當(dāng)時(shí)，一般取。當(dāng)時(shí),按式（3-3）計(jì)算 (該設(shè)計(jì)采用無(wú)縫鋼管) 式（3-3）根據(jù)缸徑查手冊(cè)預(yù)取=5此時(shí)最高允許壓力一般是額定壓力的1.5倍，根據(jù)給定參數(shù)，所以： =161.5=24MP=100110（無(wú)縫鋼管），取=100，其壁厚按公式（3-3）計(jì)算為 滿足要求，就取壁厚為5mm。5 活塞桿強(qiáng)度和液壓缸穩(wěn)定性計(jì)算A.活塞桿強(qiáng)度計(jì)算活塞桿的直徑按下式進(jìn)行校核式中，為活塞桿上的作用力；為活塞桿材料的許用應(yīng)力，=,n一般取1.40。滿足要求B.液壓缸穩(wěn)定性計(jì)算活塞桿受軸向壓縮負(fù)載時(shí)，它所承受的力不能超過(guò)使它保持穩(wěn)定工作所允許的臨界負(fù)載，以免發(fā)生縱向彎曲，破壞液壓缸的正常工作。的值與活塞桿材料性質(zhì)、截面形狀、直徑和長(zhǎng)度以及液壓缸的安裝方式等因素有關(guān)。若活塞桿的長(zhǎng)徑比且桿件承受壓負(fù)載時(shí)，則必須進(jìn)行液壓缸穩(wěn)定性校核?；钊麠U穩(wěn)定性的校核依下式進(jìn)行式中，為安全系數(shù)，一般取=24。 a.當(dāng)活塞桿的細(xì)長(zhǎng)比時(shí) b.當(dāng)活塞桿的細(xì)長(zhǎng)比時(shí)式中，為安裝長(zhǎng)度，其值與安裝方式有關(guān)，見(jiàn)表1；為活塞桿橫截面最小回轉(zhuǎn)半徑，；為柔性系數(shù)，其值見(jiàn)表3-2； 為由液壓缸支撐方式?jīng)Q定的末端系數(shù)，其值見(jiàn)表1；為活塞桿材料的彈性模量，對(duì)鋼?。粸榛钊麠U橫截面慣性矩；為活塞桿橫截面積；為由材料強(qiáng)度決定的實(shí)驗(yàn)值，為系數(shù)，具體數(shù)值見(jiàn)表3-3。表3-2液壓缸支承方式和末端系數(shù)的值支承方式支承說(shuō)明末端系數(shù)一端自由一端固定1/4兩端鉸接1一端鉸接一端固定2兩端固定4表3-3 、的值材料鑄鐵5.61/160080鍛鐵2.51/9000110鋼4.91/500085c.當(dāng)時(shí),缸已經(jīng)足夠穩(wěn)定，不需要進(jìn)行校核。此設(shè)計(jì)安裝方式中間固定的方式，此缸已經(jīng)足夠穩(wěn)定，不需要進(jìn)行穩(wěn)定性校核。6缸筒壁厚的驗(yàn)算下面從以下三個(gè)方面進(jìn)行缸筒壁厚的驗(yàn)算： A液壓缸的額定壓力值應(yīng)低于一定的極限值，保證工作安全： 式（3-4）根據(jù)式（3-4）得到：顯然，額定油壓=16MP，滿足條件；B為了避免缸筒在工作時(shí)發(fā)生塑性變形，液壓缸的額定壓力值應(yīng)與塑性變形壓力有一定的比例范圍： 式（3-5） 式（3-6）先根據(jù)式（3-6）得到：=41.21顯然，滿足條件；C耐壓試驗(yàn)壓力，是液壓缸在檢查質(zhì)量時(shí)需承受的試驗(yàn)壓力。在規(guī)定的時(shí)間內(nèi)，液壓缸在此壓力 下，全部零件不得有破壞或永久變形等異?，F(xiàn)象。各國(guó)規(guī)范多數(shù)規(guī)定: 當(dāng)額定壓力時(shí)（MPa）D為了確保液壓缸安全的使用，缸筒的爆裂壓力應(yīng)大于耐壓試驗(yàn)壓力： (MPa） 式（3-7）因?yàn)椴楸硪阎?596MPa，根據(jù)式（3-7）得到：至于耐壓試驗(yàn)壓力應(yīng)為：因?yàn)楸褖毫h(yuǎn)大于耐壓試驗(yàn)壓力，所以完全滿足條件。以上所用公式中各量的意義解釋如下：式中: 缸筒內(nèi)徑（）； 缸筒外徑（）； 液壓缸的額定壓力（） 液壓缸發(fā)生完全塑形變形的壓力（）； 液壓缸耐壓試驗(yàn)壓力（）； 缸筒發(fā)生爆破時(shí)壓力（）； 缸筒材料抗拉強(qiáng)度（）； 缸筒材料的屈服強(qiáng)度（； 缸筒材料的彈性模量（）； 缸筒材料的泊桑系數(shù) 鋼材：=0.3 7缸筒的加工要求缸筒內(nèi)徑采用H7級(jí)配合，表面粗糙度為0.16，需要進(jìn)行研磨；熱處理：調(diào)制，HB240；缸筒內(nèi)徑的圓度、錐度、圓柱度不大于內(nèi)徑公差之半；剛通直線度不大于0.03mm；油口的孔口及排氣口必須有倒角，不能有飛邊、毛刺；在缸內(nèi)表面鍍鉻，外表面刷防腐油漆。8法蘭設(shè)計(jì)液壓缸的端蓋形式有很多，較為常見(jiàn)的是法蘭式端蓋。本次設(shè)計(jì)選擇法蘭式端蓋（缸筒端部）法蘭厚度根據(jù)下式進(jìn)行計(jì)算： 式（3-8）式中， -法蘭厚度（m）；密封環(huán)內(nèi)經(jīng)d=40mm（m）；密封環(huán)外徑（m）；=50mm系統(tǒng)工作壓力（pa）；=7MPa附加密封力（Pa）；值取其材料屈服點(diǎn)353MPa；螺釘孔分布圓直徑（m）；=55mm密封環(huán)平均直徑（m）；=45mm法蘭材料的許用應(yīng)力（Pa）；=/n=353/5=70.6MPa法蘭受力總合力（m） 所以=13.2mm為了安全取=14mm9缸筒端部）法蘭連接螺栓的強(qiáng)度計(jì)算連接圖如下：圖3-1缸體端部法蘭用螺栓連接1-前端蓋；2-缸筒螺栓強(qiáng)度根據(jù)下式計(jì)算：螺紋處的拉應(yīng)力:(MPa) 式（3-9）螺紋處的剪應(yīng)力(MPa) 式（3-10）合成應(yīng)力 (MPa) 式（3-11）式中, 液壓缸的最大負(fù)載，=A，單桿時(shí)，雙桿是螺紋預(yù)緊系數(shù)，不變載荷=1.251.5，變載荷=2.54；液壓缸內(nèi)徑；缸體螺紋外徑；螺紋內(nèi)經(jīng)；螺紋內(nèi)摩擦因數(shù)，一般取=0.12;變載荷取=2.54；材料許用應(yīng)力，,為材料的屈服極限，n為安全系數(shù)，一般取n=1.21.5；Z螺栓個(gè)數(shù)。最大推力為：使用4個(gè)螺栓緊固缸蓋，即：=4螺紋外徑和底徑的選擇：=10mm =8mm系數(shù)選擇：選取=1.3=0.12根據(jù)式（3-9）得到螺紋處的拉應(yīng)力為：=根據(jù)式（3-10）得到螺紋處的剪應(yīng)力為：根據(jù)式（3-11）得到合成應(yīng)力為：=367.6MPa由以上運(yùn)算結(jié)果知，應(yīng)選擇螺栓等級(jí)為12.9級(jí)；查表的得：抗拉強(qiáng)度極限=1220MP;屈服極限強(qiáng)度=1100MP；不妨取安全系數(shù)n=2可以得到許用應(yīng)力值：=/n=1100/2=550MP證明選用螺栓等級(jí)合適。 10密封件的選用A.對(duì)密封件的要求在液壓元件中，液壓缸的密封要求是比較高的，特別是一些特殊液壓缸，如擺動(dòng)液壓缸等。液壓缸不僅有靜密封，更多的部位是動(dòng)密封，而且工作壓力高，這就要求密封件的密封性能要好，耐磨損，對(duì)溫度的適應(yīng)范圍大，要求彈性好，永久變形小，有適當(dāng)?shù)臋C(jī)械強(qiáng)度，摩擦阻力小，容易制造和裝拆，能隨壓力的升高而提高密封能力和利于自動(dòng)補(bǔ)償磨損。密封件一般以斷面形狀分類，有O形、Y形、U形、V形和Yx形等。除O形外，其他都屬于唇形密封件。B. O形密封圈的選用液壓缸的靜密封部位主要有活塞內(nèi)孔與活塞桿、支撐座外圓與缸筒內(nèi)孔、端蓋與缸體端面等處。靜密封部位使用的密封件基本上都是O形密封圈。C.動(dòng)密封部位密封圈的選用由于O型密封圈用于往復(fù)運(yùn)動(dòng)存在起動(dòng)阻力大的缺點(diǎn)，所以用于往復(fù)運(yùn)動(dòng)的密封件一般不用O形圈，而使用唇形密封圈或金屬密封圈。液壓缸動(dòng)密封部位主要有活塞與缸筒內(nèi)孔的密封、活塞桿與支撐座（或?qū)蛱祝┑拿芊獾?。活塞環(huán)是具有彈性的金屬密封圈，摩擦阻力小，耐高溫，使用壽命長(zhǎng)，但密封性能差，內(nèi)泄漏量大，而且工藝復(fù)雜，造價(jià)高。對(duì)內(nèi)泄漏量要求不嚴(yán)而要求耐高溫的液壓缸，使用這種密封圈較合適。V形圈的密封效果一般，密封壓力通過(guò)壓圈可以調(diào)節(jié)，但摩擦阻力大，溫升嚴(yán)重。因其是成組使用，模具多，也不經(jīng)濟(jì)。對(duì)于運(yùn)動(dòng)速度不高、出力大的大直徑液壓缸，用這種密封圈較好。U形圈雖是唇形密封圈，但安裝時(shí)需用支撐環(huán)壓住，否則就容易卷唇，而且只能在工作壓力低于10MPa時(shí)使用，對(duì)壓力高的液壓缸不適用。比較而言，能保證密封效果，摩擦阻力小，安裝方便，制造簡(jiǎn)單經(jīng)濟(jì)的密封圈就屬Yx型密封圈了。它屬于不等高雙唇自封壓緊式密封圈 ，分軸用和孔用兩種。綜上，所以本設(shè)計(jì)選用Yx型圈，聚氨酯和聚四氟乙烯密封材料組合使用，可以顯著提高密封性能：a.降低摩擦阻力，無(wú)爬行現(xiàn)象；b.具有良好的動(dòng)態(tài)和靜態(tài)密封性，耐磨損，使用壽命長(zhǎng)；c.安裝溝槽簡(jiǎn)單，拆裝簡(jiǎn)便。這種組合的特別之處就是允許活塞外園和缸筒內(nèi)壁有較大間隙，因?yàn)榻M合式密封的密封圈能防止擠入間隙內(nèi)，降低了活塞與缸筒的加工要求，密封方式圖如下：圖3-2 密封方式圖3.5 制定液壓回路方案，擬定液壓系統(tǒng)原理圖擬定液壓系統(tǒng)圖 圖3-2液壓專用鉆床液壓系統(tǒng)原理圖1主液壓泵 ； 2定量泵 ；3、4溢流閥 ；5遠(yuǎn)程調(diào)壓閥 ；6電液換向閥 ；7壓力表 ；8電磁換向閥；9液控單向閥 ；10順序閥 ；11卸荷閥（帶阻尼孔） ；12壓力繼電器 ；13單向閥 ；14充液閥 ；15充液箱 ；16液壓缸； 17滑塊；18擋鐵。3.6液壓系統(tǒng)的工作原理液壓缸的工作循環(huán)為：快進(jìn)工進(jìn)死擋鐵停留快退停止。1)快速下行電磁鐵2DT和3DT通電，電液換向閥6和電磁換向閥8均換至右工位，后者使液控單向閥9打開(kāi)。此時(shí)液壓缸進(jìn)回液路區(qū)暢通。進(jìn)油路：主液壓泵1 電液換向閥6 單向閥13 液壓缸18上（無(wú)桿）腔；回油路：液壓缸18下（有桿）腔 液控單向閥9 電液換向閥6 油箱。此時(shí)液壓缸滑塊16因自重而快速下降，主液壓泵1全部流量尚不能滿足快速要求的流量，液壓缸18上腔形成局部真空，呈泵工況，油箱（置于液壓缸頂部）中油液在大氣壓力下經(jīng)液控充液閥（液控單向閥）14充入，避免了上述不利現(xiàn)象產(chǎn)生。2)慢速接近工件和逐步加壓擋鐵17壓下行程開(kāi)關(guān)XK2時(shí)，電磁鐵3DT斷電，電磁換向閥8處于常態(tài)（圖示位置），液控單向閥9關(guān)閉，閥芯緊閉。進(jìn)油路：主液壓泵1 電液換向閥6 單向閥13 液壓缸18上腔；回油路：液壓缸18下腔 順序閥10 電液換向閥6 油箱。順序閥10使下腔建立起背壓，滑塊靠自重不能下降，主液壓泵1供給的壓力油使之下行。這時(shí)上腔壓力升高，充液閥（液控單向閥）14關(guān)閉，活塞速度降低。當(dāng)滑塊慢速接觸工件時(shí)，阻力（負(fù)載）急劇增加，主液壓泵1工作壓力急劇升高，排量自動(dòng)減小，液壓缸活塞速度進(jìn)一步降低，以極慢的速度對(duì)工件加壓。3)保壓延時(shí)當(dāng)液壓缸18工作壓力達(dá)到預(yù)定值時(shí)，壓力繼電器12發(fā)出電氣控制信號(hào)，電磁鐵2DT斷電，電液換向閥6復(fù)中位，液壓缸進(jìn)回液腔封閉，主液壓泵1經(jīng)電液換向閥6中位卸荷。保壓時(shí)間可由壓力繼電器12控制的時(shí)間繼電器調(diào)節(jié)。4)快速回程保壓結(jié)束后，時(shí)間繼電器發(fā)出信號(hào)使電磁鐵2DT斷電，1DT通電，電液換向閥6切至左位，同時(shí)進(jìn)油路控制油液使充液閥（液控單向閥）14打開(kāi)，為液壓缸18退回做好準(zhǔn)備。這時(shí)：進(jìn)油路：主液壓泵1 電液換向閥6 液控單向閥9 液壓缸18下腔；回油路：液壓缸18上腔 充液閥（液控單向閥）14 油箱。需要說(shuō)明的是，電液換向閥6切至左位時(shí)，液壓缸18還未泄壓時(shí)，上腔壓力很高，卸荷閥11（帶阻尼孔）呈開(kāi)放狀態(tài)，主液壓泵1的輸出油液經(jīng)此閥阻尼孔回油箱，這時(shí)主液壓泵1工作壓力較低，不足以使液壓缸回程，但可使充液閥（液控單向閥）14開(kāi)啟，使液壓缸18上腔泄壓；當(dāng)液壓缸上腔壓力降到定值時(shí)，卸荷閥11關(guān)閉，此時(shí)主液壓泵1才開(kāi)始向液壓缸18下腔供液，液壓缸快速回程。5)停止液壓缸位于其反向行程末端時(shí)，擋鐵下壓行程開(kāi)關(guān)XK1，電磁鐵1DT斷電，電液換向閥6處于中位，液壓缸被鎖而停止。主液壓泵1此時(shí)處于卸荷狀態(tài)。在使用中，可隨時(shí)手動(dòng)控制1DT斷電，使液壓缸隨時(shí)處于停止?fàn)顟B(tài)。其工作循環(huán)和電磁鐵動(dòng)作順序表如表3-1所示。表3-1液壓專用鉆床工作循環(huán)和電磁鐵動(dòng)作順序表動(dòng)作名稱信號(hào)來(lái)源換向滑閥工作狀態(tài)電磁鐵動(dòng)態(tài)狀態(tài)電液換向閥6電磁換向閥81DT2DT3DT液壓缸快速下行2DT和3DT通電右位右位+慢速加壓擋鐵行程開(kāi)關(guān)XK2，3DT斷電，4DT通電右位常態(tài)+保壓延時(shí)壓力繼電器12發(fā)出信號(hào)，2DT斷電中位快速回程壓力繼電器12發(fā)出信號(hào)，1DT通電左位+停止行程開(kāi)關(guān)XK1發(fā)出信號(hào)，1DT斷電中位3.7 計(jì)算與選擇液壓元件 3.7.1 液壓泵及驅(qū)動(dòng)電機(jī)計(jì)算與選定 （1）、液壓泵的選擇液壓泵的最高工作壓力計(jì)算 由工況圖4-1可以查得液壓缸的最高工作壓力出現(xiàn)在工進(jìn)階段，即由于進(jìn)油路元件較少，故泵至缸間的進(jìn)油路壓力損失估取為。則液壓泵的最高工作壓力為 所需的液壓泵最大供油量qp按液壓缸的最大輸入流量估算。取泄漏系數(shù)K=1.1則 qp=1.1* 18.4=20.24(L/min) 暫取泵的容積效率v=0.90可算得泵的排量參考值為 Vg=1000qv/nv=1000*20.24/1500*0.9=14.9mL/r 根據(jù)以上計(jì)算結(jié)果查閱產(chǎn)品樣本，選用規(guī)格相近的25YCY141B壓力補(bǔ)償變量型斜盤(pán)式軸向柱塞泵，其額定壓力Pn=32Mpa，V=25mL/r，n=1500r/min，容積效率v=0.92，qp=Vnv=25*1500*0.92=34.5L/min，符合系統(tǒng)對(duì)流量的要求（2）、電動(dòng)機(jī)的選擇 固定設(shè)備的液壓系統(tǒng)，其液壓泵通常用電動(dòng)機(jī)驅(qū)動(dòng)。 根據(jù)算出的功率和液壓泵的轉(zhuǎn)速及其使用環(huán)境，從產(chǎn)品樣本或手冊(cè)中選定其型號(hào)規(guī)格額定功率、轉(zhuǎn)速、電源、結(jié)構(gòu)形式（立式、臥式，開(kāi)式、封閉式的等），并對(duì)其進(jìn)行核算，以保證每個(gè)工作階段電動(dòng)機(jī)的峰值超載量都低于25%。 由于液壓泵通常在空載下啟動(dòng)，故對(duì)電動(dòng)機(jī)的啟動(dòng)轉(zhuǎn)矩沒(méi)有過(guò)高的要求，負(fù)荷變化比較平穩(wěn)，啟動(dòng)次數(shù)不多，故可采用籠型三相異步電動(dòng)機(jī)。但若液壓系統(tǒng)功率較大而電網(wǎng)容量不大時(shí)，可采用繞線轉(zhuǎn)子電動(dòng)機(jī)。對(duì)于采用變頻調(diào)節(jié)流量方案的液壓泵，則應(yīng)采用變頻調(diào)速或電磁調(diào)速控制的交流異步電動(dòng)機(jī)驅(qū)動(dòng)液壓泵。 由工況圖知，最大功率出現(xiàn)在終壓階段t=0.395s時(shí)，由此時(shí)的液壓缸工作壓力和流量可算得此時(shí)液壓泵的最大理論功率 Pt=（p+p）Kq=（8+0.5）*（1.1*4.7）/60=0.73Kw取泵的總效率為p=0.85，則算得液壓泵驅(qū)動(dòng)功率為 Pp=Pt/p=0.73/0.85=0.86Kw查手冊(cè)，選用規(guī)格相近的Y90L14型封閉式三相異步電動(dòng)機(jī)，轉(zhuǎn)速1440r/min，額定功率為1.5Kw。 按所選電動(dòng)機(jī)轉(zhuǎn)速和液壓泵的排量，液壓泵的最大實(shí)際流量為 大于計(jì)算所需流量20.24L/min,滿足使用要求。3.7.2 液壓控制閥和液壓輔助元件的選定 根據(jù)所選擇的液壓泵規(guī)格和系統(tǒng)的工作情況，容易選擇系統(tǒng)的其他液壓元件，一并列入表8-1序號(hào)元件名稱估計(jì)通過(guò)流量型號(hào)規(guī)格1斜盤(pán)式柱塞泵2525YCY141B32Mpa，驅(qū)動(dòng)功率24.6KN2WU網(wǎng)式濾油器25WU-25*18015通徑，壓力損失0.01MPa3直動(dòng)式溢流閥12YEF-10B10通徑，32Mpa，板式聯(lián)接4背壓閥63YF3-10B10通徑，21Mpa，板式聯(lián)接5二位二通手動(dòng)電磁閥8022EF3-E10B6三位四通電磁閥 6034F3-Ea6B6通徑，壓力31.5MPa7液控單向閥40YAF3-Ea10B32通徑，32MPa8調(diào)速閥80QFF3-E10B10通徑，16MPa9調(diào)速閥80QF3-E10B10通徑，16MPa10二位二通電磁閥3022EF3B-E10B6通徑，壓力20 MPa11壓力繼電器DP1-63B8通徑，10.5-35 MPa12壓力表開(kāi)關(guān)KF3-E3B32Mpa，6測(cè)點(diǎn)13油箱14液控單向閥YAF3-Ea10B32通徑，32MPa15上液壓缸16下液壓缸17單向節(jié)流閥48ALF3E10B10通徑，16MPa18單向單向閥48ALF3E10B10通徑，16MPa19三位四通電磁換向閥2534EF30-E6B6通徑，16MPa20減壓閥40JF3-10B10通徑，板式連接3.7.3油管的選擇油管系統(tǒng)中使用的油管種類很多，有鋼管、銅管、尼龍管、塑料管、橡膠管等，必須按照安裝位置、工作環(huán)境和工作壓力來(lái)正確選用。本設(shè)計(jì)中油管采用鋼管，因?yàn)楸驹O(shè)計(jì)中所須的壓力是高壓，P=16MPa ， 鋼管能承受高壓，價(jià)格低廉，耐油，抗腐蝕，剛性好，但裝配是不能任意彎曲，常在裝拆方便處用作壓力管道一中、高壓用無(wú)縫管，低壓用焊接管。本設(shè)計(jì)在彎曲的地方可以用管接頭來(lái)實(shí)現(xiàn)彎曲。尼龍管用在低壓系統(tǒng)；塑料管一般用在回油管用。膠管用做聯(lián)接兩個(gè)相對(duì)運(yùn)動(dòng)部件之間的管道。膠管分高、低壓兩種。高壓膠管是鋼絲編織體為骨架或鋼絲纏繞體為骨架的膠管，可用于壓力較高的油路中。低壓膠管是麻絲或棉絲編織體為骨架的膠管，多用于壓力較低的油路中。由于膠管制造比較困難，成本很高，因此非必要時(shí)一般不用。1. 管接頭的選用：管接頭是油管與油管、油管與液壓件之間的可拆式聯(lián)接件，它必須具有裝拆方便、連接牢固、密封可靠、外形尺寸小、通流能力大、壓降小、工藝性好等各種條件。管接頭的種類很多，液壓系統(tǒng)中油管與管接頭的常見(jiàn)聯(lián)接方式有：焊接式管接頭、卡套式管接頭、擴(kuò)口式管接頭、扣壓式管接頭、固定鉸接管接頭。管路旋入端用的連接螺紋采用國(guó)際標(biāo)準(zhǔn)米制錐螺紋（ZM）和普通細(xì)牙螺紋（M）。錐螺紋依靠自身的錐體旋緊和采用聚四氟乙烯等進(jìn)行密封，廣泛用于中、低壓液壓系統(tǒng)；細(xì)牙螺紋密封性好，常用于高壓系統(tǒng)，但要求采用組合墊圈或O形圈進(jìn)行端面密封，有時(shí)也采用紫銅墊圈。2.管道內(nèi)徑計(jì)算： （1）式中 Q通過(guò)管道內(nèi)的流量 v管內(nèi)允許流速 ，見(jiàn)表：表3.2：液壓系統(tǒng)各管道流速推薦值油液流經(jīng)的管道推薦流速 m/s液壓泵吸油管0.51.5液壓系統(tǒng)壓油管道36，壓力高，管道短粘度小取大值液壓系統(tǒng)回油管道1.52.6 (1). 液壓泵壓油管道的內(nèi)徑： 取v=4m/s 根據(jù)簡(jiǎn)明手冊(cè)P111查得：取d=20mm,鋼管的外徑 D=28mm; 管接頭聯(lián)接螺紋M272。(2) . 液壓泵回油管道的內(nèi)徑：取v=2.4m/sd=21mm根據(jù)簡(jiǎn)明手冊(cè)P111查得：取d=25mm,鋼管的外徑 D=34mm; 管接頭聯(lián)接螺紋M332。3. 管道壁厚的計(jì)算 式中： p管道內(nèi)最高工作壓力 Pa d管道內(nèi)徑 m管道材料的許用應(yīng)力 Pa，管道材料的抗拉強(qiáng)度 Pan安全系數(shù)，對(duì)鋼管來(lái)說(shuō)，時(shí)，取n=8；時(shí)，取n=6； 時(shí)，取n=4。根據(jù)上述的參數(shù)可以得到：我們選鋼管的材料為45#鋼，由此可得材料的抗拉強(qiáng)度=600MPa; (1). 液壓泵壓油管道的壁厚(2). 液壓泵回油管道的壁厚3.7.4液壓系統(tǒng)的驗(yàn)算前述液壓系統(tǒng)的初步設(shè)計(jì)是在某些估計(jì)參數(shù)的情況下進(jìn)行的，當(dāng)液壓系統(tǒng)原理圖，組成元件及連接管路等完全確定后，針對(duì)實(shí)際情況對(duì)設(shè)計(jì)的系統(tǒng)進(jìn)行各項(xiàng)性能分析計(jì)算，其目的在于對(duì)液壓系統(tǒng)的設(shè)計(jì)質(zhì)量作出評(píng)價(jià)和評(píng)判，若出現(xiàn)問(wèn)題，則應(yīng)對(duì)液壓系統(tǒng)某些不合理的設(shè)計(jì)進(jìn)行修正或重新調(diào)整，或采取其他的必要的措施，性能驗(yàn)算內(nèi)容一般包括壓力損失，效率，發(fā)熱與升溫，液壓沖擊等，對(duì)于較重要的系統(tǒng)，還應(yīng)對(duì)其動(dòng)態(tài)性能進(jìn)行驗(yàn)算或計(jì)算機(jī)仿真。計(jì)算時(shí)通常只采用一些簡(jiǎn)化公式以求得概略結(jié)果。1、液壓系統(tǒng)壓力損失驗(yàn)算上面已經(jīng)計(jì)算出該液壓系統(tǒng)中進(jìn)，回油管的內(nèi)徑分別為20mm,25mm。但是由于系統(tǒng)的具體管路布置和長(zhǎng)度尚未確定，所以壓力損失無(wú)法驗(yàn)算。（1）工作進(jìn)給時(shí)進(jìn)油路壓力損失。運(yùn)動(dòng)部件工作進(jìn)給時(shí)的最大速度為50mm/s，進(jìn)給時(shí)的最大流量為18.7L/min，則液壓油在管內(nèi)流速為 管道雷諾數(shù)為，可見(jiàn)油液在管道內(nèi)流態(tài)為層流，其沿程阻力系數(shù)進(jìn)油管道BC的沿程壓力損失為查得換向閥34EF30-E6B的壓力損失忽略油液通過(guò)管接頭、油路板等處的局部壓力損失，則進(jìn)油路總壓力損失為（2）工作進(jìn)給時(shí)回油路的壓力損失。由于選用單活塞桿液壓缸，且液壓缸有桿腔的工作面積為無(wú)桿腔的工作面積的二分之一，則回油管道的流量為進(jìn)油管道的二分之一，則回油管道的沿程壓力損失為：查產(chǎn)品樣本知換向閥23EF3B-E10B的壓力損失，換向閥34EF30-E10B的壓力損失，調(diào)速閥AQF3-E10B壓力損失?；赜吐房倝毫p失為（3）變量泵出口處得壓力（4）快進(jìn)時(shí)的壓力損失。快進(jìn)時(shí)液壓缸為差動(dòng)連接，自匯流點(diǎn)A至液壓缸進(jìn)油口C之間的管路AC中，流量為液壓泵出口流量的兩倍即390L/min，AC段管路的沿程壓力損失為同樣可求管道AB段及AD段得沿程壓力損失和為查產(chǎn)品樣本知，流經(jīng)各閥的局部壓力損失為：34EF30-E10B的壓力損失，23EF3B-E10B的壓力損失據(jù)分析在差動(dòng)連接中，泵的出口壓力為快退時(shí)壓力損失驗(yàn)算從略。上述驗(yàn)算表明，無(wú)需修改原設(shè)計(jì)。2、液壓系統(tǒng)效率的估算 估算液壓系統(tǒng)效率時(shí)，主要應(yīng)考慮液壓泵的總效率p、液壓執(zhí)行元件的總效率A及液壓回路的效率C。=PC3、系統(tǒng)溫升的驗(yàn)算在整個(gè)工作循環(huán)中，工進(jìn)階段所占的時(shí)間最長(zhǎng)，且發(fā)熱量最大。為了簡(jiǎn)化計(jì)算，主要考慮工進(jìn)時(shí)的發(fā)熱量。一般情況下，工進(jìn)時(shí)做功的功率損失大引起發(fā)熱量較大，所以只考慮工進(jìn)時(shí)的發(fā)熱量，然后取其值進(jìn)行分析。當(dāng)V=10mm/s時(shí)，即v=600mm/min 即q=7.4L/min此時(shí)泵的效率為0.9，泵的出口壓力為20MP，則有kw此時(shí)的功率損失為：假定系統(tǒng)的散熱狀況一般，取，油箱的散熱面積A為系統(tǒng)的溫升為油箱中溫度一般推薦30-50所以驗(yàn)算表明系統(tǒng)的溫升在許可范圍內(nèi)。4 設(shè)計(jì)小結(jié)隨著大學(xué)四年生活的即將結(jié)束，我們業(yè)即將踏上建設(shè)祖國(guó)的征途。大學(xué)四年生活的點(diǎn)點(diǎn)滴滴都斗匯聚到這幾個(gè)月。經(jīng)過(guò)幾個(gè)月的苦戰(zhàn)我的課程設(shè)計(jì)終于要完成了。在以前我們也做過(guò)設(shè)計(jì)，所以也就認(rèn)為課程設(shè)計(jì)沒(méi)有什么難的，只是對(duì)以前所學(xué)的知識(shí)的檢驗(yàn)。但是真正做過(guò)了這次課程設(shè)計(jì)以后我才發(fā)現(xiàn)原來(lái)我們以前做的并不能叫做設(shè)計(jì)，至少不那么規(guī)范。而對(duì)于即將作為社會(huì)的主人的我們這一點(diǎn)是必須的。以前老是覺(jué)得自己什么東西都會(huì)，什么東西都懂，有點(diǎn)眼高手低。通過(guò)這次課程設(shè)計(jì)，我才明白原來(lái)學(xué)習(xí)是一個(gè)長(zhǎng)期積累的過(guò)程，在以后的工作、生活中都應(yīng)該不斷的學(xué)習(xí)，努力提高自己知識(shí)和綜合素質(zhì)。課程設(shè)計(jì)是我們專業(yè)課程知識(shí)綜合應(yīng)用的訓(xùn)練，是我們邁向社會(huì)，從事職業(yè)工作前一個(gè)必不少的過(guò)程。”千里之行，始于足下”,通過(guò)這次課程設(shè)計(jì)，我深深體會(huì)到這句千古名言的真正含義，我今天認(rèn)真的進(jìn)行課程設(shè)計(jì)，學(xué)會(huì)腳踏實(shí)地的邁開(kāi)這一步，就是為明天能夠穩(wěn)健的在社會(huì)大潮奔跑打下堅(jiān)實(shí)的基礎(chǔ)。在這次設(shè)計(jì)過(guò)程中，體現(xiàn)出自己?jiǎn)为?dú)設(shè)計(jì)的能力以及綜合運(yùn)用知識(shí)的能力，體會(huì)了學(xué)以致用、突出自己勞動(dòng)成果的喜悅心情，從中發(fā)現(xiàn)自己平時(shí)學(xué)習(xí)的不足和薄弱環(huán)節(jié)，從而加以彌補(bǔ)。經(jīng)過(guò)了這么多的過(guò)程一個(gè)合格的設(shè)計(jì)過(guò)程才能交完成。通過(guò)這次設(shè)計(jì)我學(xué)到了很多東西，這不僅僅是知識(shí)層面上的，我們?cè)谧鋈松弦惨徊揭粋€(gè)腳印，踏踏實(shí)實(shí)的。最后預(yù)祝我們?cè)趯?lái)的人生中做一個(gè)合格的人。致 謝首先，我要感謝學(xué)校這四年來(lái)的培養(yǎng)，讓我有機(jī)會(huì)在更深的層次上進(jìn)行學(xué)習(xí)。同時(shí)我要感謝四年來(lái)各位老師朋友對(duì)我的關(guān)心指導(dǎo)。在此我要特別感謝老師在課程設(shè)計(jì)中對(duì)我的悉心指導(dǎo)。在論文研究和撰寫(xiě)過(guò)程中，本人得到了指導(dǎo)教師老師的悉心指導(dǎo)。在課題研究過(guò)程中，老師淵博的學(xué)識(shí)、踏實(shí)嚴(yán)謹(jǐn)?shù)闹螌W(xué)態(tài)度以及精益求精的工作作風(fēng)使我獲益匪淺。在生活中，老師認(rèn)真負(fù)責(zé)的良師風(fēng)范使我懂得了很多為人的道理。在此，謹(jǐn)向老師表示最衷心的感謝！感謝和我同組的各位同學(xué)，我們一起探討問(wèn)題，一起解決問(wèn)題，在整個(gè)設(shè)計(jì)過(guò)程中，他們給了我很大的鼓勵(lì)和支持。感謝我的室友們，我們來(lái)自四面八方，是你們的友情支撐我走過(guò)了這漫長(zhǎng)的四年。四年了，仿佛就在昨天。四年里，我們彼此間充滿歡笑。只是，只是我們以后也許在難相見(jiàn)了，沒(méi)關(guān)系，預(yù)祝大家前程似錦，珍重。我會(huì)記住我們?cè)谝黄鸬拿篮脮r(shí)光。在論文即將完成之際，我的心情無(wú)法平靜，從開(kāi)始進(jìn)入課題到論文的順利完成，得到了很多多位同學(xué)們的支持與幫助，在此一并表示感謝，我們永遠(yuǎn)是朋友，今生勿忘！參考文獻(xiàn)1張利平.液壓傳動(dòng)系統(tǒng)設(shè)計(jì)及使用. 化學(xué)工業(yè)出版社，20042楊培元. 液壓系統(tǒng)設(shè)計(jì)簡(jiǎn)明手冊(cè).北京：機(jī)械工業(yè)出版社，20083機(jī)械設(shè)計(jì)手冊(cè)編委會(huì).機(jī)械設(shè)計(jì)手冊(cè)第四卷.北京：機(jī)械工業(yè)出版社，20074王積偉.液壓與氣壓傳動(dòng).北京：北京出版社，20055張利平.液壓與氣壓技術(shù)速查手冊(cè).北京：化學(xué)工業(yè)出版社，20066張利平.現(xiàn)代液壓技術(shù)應(yīng)用220例.北京：化學(xué)工業(yè)出版社，20077王守成.液壓元件及選用.北京：化學(xué)工業(yè)出版社，20078鄧樂(lè).液壓傳動(dòng).北京：北京郵電大學(xué)出版社，20109劉延俊.液壓元件使用指南.北京：化學(xué)工業(yè)出版社，200710Anthony Esposito.Fluid Power With Applications.New Jersey:Prentice-rlass.lnc.198011Rob Walter.Improring systems with accumulaters means big saeings. Hydraulics&Pneumat -cs.Aug2008:40 KSME International Journal, Vol. 11, No. 4, pp. 397-407, 1997 A Study on the Stability Analysis of a PWM Controlled Hydraulic Equipment 397 Jun-Young Huh* and G. Wennmacher* (Received August 16 1996) PWM control, which is inherently nonlinear digital control, has been used for hydraulic equipment control because of the robustness and the availability of the low priced high speed on-off valve which is required for the system. Since this valve can be directly controlled without any D/A converter, it is easily implemented to hydraulic equipments with microcomputers. The objectives of this study is to analyze the limit cycle which ordinarily appears in the position control system using high speed on-off valve, and to give a criterion for the stability of this sy,;tem. The nonlinear characteristics of PWM and cylinder friction are described by harmonic linearization and the effects of parameter variations to the system stability are investigated theoretically. Experimental results demonstrate the feasibility of the proposed method. Key Words: Electro-Hydraulic Servo System, High Speed On-Off Solenoid Valve, PWM Method, Stability Analysis, Limit Cycle Nomenclature Ns : Transfer function of valve-cylinder sys- tem As : Piston area P1, P2 : Forward and return pressure of cylinder ca : Discharge coefficient Ps : Supply pressure D : Duty (:acting time/T) Pr : Tank pressure Fc : Coulombs friction force q : Input signal of PWM FL : External load c : Amplitude of PWM input signal Fe : Friction force of piston Q, Q2 : Forward and return flows G : Viscous damping coefficient qmax : Magnitude of PWM input signal of duty kl : Flow gain of valve : 1 k2 : Flow-pressure coefficient T : Period of PWM carrier wave k: : Proportional gain Tq : Period of PWM input signal wave km : Maximum displacement of valve poppet Ts : Period of sampling time K : System gain (:k: 9 k,) : Amplitude of PWM output pulse Kq : Sizing factor of valve v : Velocity of piston Ls : Static friction 7 : Amplitude of piston velocity wave M : Piston mass Vt : Total volume of pipe, valve and cylinder MRa : Critical speed starting viscous damping chamber friction w : Area gradient of valve port NpwM : Describing function of PWM x : Displacement of piston xv : Displacement of valve poppet (i:a, b, 9 Department of Mechanical Engineering, Korea c, d) University of Technology and Education, Chonan P.O.B. 55, Chungnam, 330-600, Korea Be Effective bulk modulus of fluid and pipe 9 * Liebherr-Aerospace Lindenberg GmbH, Pfaender : Synchronizing degree Str. 50-52, D-88161 Lindenberg, Germany r : Valve delay time 398 Jun-Young Huh and G. Wennrnacher r(q) : Width of PWM output co : Angular frequency of PWM input signal ton.ea : Switch on delay ton,d,so : Switch on displacement time tof.dead : Switch off delay toff,dsp : Switch off displacement time 1. Introduction The control method using Pulse Width Modu- lation(PWM) is one of the nonlinear control schemes. Hydraulic equipments which are operat- ed in PWM control usually use high speed on-off valves as control valves. The usage of this valve gives several advantages; this valve has very sim- ple structure so that it has robustness to oil contamination and its price is low. And it can be easily implemented to the system by using micro- computer since the control of this valve can be carried out digitally without D/A converter (Wennmacher, 1992a, b; Wennmacher, 1994; Muto, et. al., 1988). Since the hydraulic cylinder system controlled by PWM has highly nonlinear characteristics such as the nonlinearity of PWM, the phase delay of valve poppet and the friction in the cylinder, the nonlinear characteristics should be considered in the analysis of the system stabil- ity. Tanaka(1988) considered this system as a linear time invariant discrete system and derived transfer function using z-transform, then, showed the stable limiting gain against the damping ratio on every sampling time. Noritsugu(1983) sim- plified the nonlinearity of PWM as a saturation function and considered it with the phase lag of valve poppet and the cylinder friction in the velocity control of pneumatic cylinder. But the nonlinearity of PWM signal generation is not considered fully. Prochnio (1986) utilized describ- ing function based on the harmonic linearization in depicting the nonlinearity of PWM and consid- ered the friction in cylinder, did not include the phase lag of the valve poppet. The objective of this study is to present a method to predict the stability of hydraulic servo system. The electro-hydraulic position control system is theoretically modelled, which have high speed on-off valves controlled by PWM. In this process, the highly nonlinear characteristics of the system such as the dynamics of PWM, the phase lag of valve poppet and the cylinder friction are modelled mathematically. Especially the non- linear characteristics of PWM and cylinder fric- tion are expressed with describing function by the harmonic linearization. The validity of the presented method is investigated for the variation of the system parameters theoretically and experi- mentally. 2. Hydraulic Servo System Description and Analysis 2.1 Dynamic modelling The schematic diagram of the hydraulic servo system employed in this study is shown in Fig. 1. This system consists of 4 high speed on-off valves and a hydraulic actuator(double rod double acting cylinder) as the main parts. The piston displacement (x), which is output of the system is feedbacked to the comparator, which is im- plemented by software inside the computer, to be compared with the reference input value. For the PWM operation with the high speed on-off valve, since the switching characteristic of valve poppet makes a significant influence to the performance of this system, the lag of the valve poppet move- ment should be considered in dynamic modelling. The hydraulic system is modelled in Fig. 2. When the ON signal is applied to the valves a and c, since the valves b and d receive the OFF signal simultaneously, the forward oil flow(Q1) is oc- IIII Fig. 1 ? Electro-hydraulic servo system with high speed on-off valves. A Study on the Stability Analysis of a PWM Controlled Hydraulic Equipment 399 X,= X,)%,R P1 P2 . l I 7- / I 7 TM 5 r I ELI Xvd Xvc I Xvb Xvo I I PT I PT PS Fig. 2 Modelling diagram of hydraulic system. ing, because it is generally not critical or of particular interest in system design, it cannot be altered appreciably by design, and it does not affect system stability. To derive the transfer func- tion of the system, the load flow Eq. in (5) which is a nonlinear equation is linealized by the Taylors expansion in the vicinity of an operating point(xv0, PLo) of interest, It can be written as (Merrit, 1967) QL = kx - k2PL (8) curred through the valves a and c, and this makes the positive piston displacement. Under the assumption that the supply pressure(Ps) is con- stant, the leakage flows between a valve poppet and its seat and that of cylinder are negligible, and the valves are operated simultaneously, the flow rates (Q, Qz) through the orifices of the high speed on-off valves are given by Q = caw v 9 xe (1) Q2=cawvf. xaPUA=-P (2) where the load pressure(PL) and the load flow (QL) are defined as pL = p- p2 (3) QL :- Q1 + Qz (4) 2 The load flow equation can be deriven as, QL = Kqx,/-Ps- PL (5) where Kq - CdW v The continuity equation of the valve-cylinder system is given by QL=AhI lit dPL (6) 4fie dt The force balance equation of the piston is d2x M-t- = AhPL - FR (7) where the viscous damping friction of the piston is included in the term of total friction force of cylinder (FR), and the external load (FL) is omit- ted. The external force is neglected in the modell- 2.2 Describing function of PWM When a high speed on-off valve is operated in PWM mode, the input signal which is inflicted on the valve from the PWM module is a type of pulse row which has a constant pulse magnitude but a different pulse width and sign according to the magnitude of the input signal of PWM module. And the pulse row has the same interval between the pulse starting points. To investigate the stabil- ity of this system analytically, the characteristics of the PWM signal generation should be depicted analytically. Under the assumption that the on- off operations of the valves are symmetric and the linear part of this hydraulic system has the enough low pass filter function (Follinger, 1991), we derive the describing function of PWM signal generation which has the highly nonlinear charac- teristic. The PWM module catches the input sig- nal(q) at every time interval and knows the end of the pulse width previously. The input signal (q (t) of the PWM module can be represented as the following in the harmonic vibration balance state. q (kT) = c7 9 sin (w 9 kT + q5) (9) where the synchronizing degree (if) represents the phase lag between the first output pulse of PWM module and the input signal of PWM module in the harmonic vibration balance state. For exam- ple, the case of if=90 is shown in Fig. 3. The input signal q(t) of PWM module which is depicted as Eq. (9) is a periodic signal which have a period (Tq). If the ratio between the period (Tq) and the period of PWM carrier wave(T) comes to 2n (n=l, 2, 3, .), Fourier series method can be utilized in calculation of the 400 Jun-Young Huh and G. Wennmacher describing function of the PWM signal genera- tion. The output pulse equation of PWM module u (t) = 0 sgn(q (kT) ) is given by kT Nt kT +rq(kT) kT+rq(kT)t (k+l) T (lO) where rq(kT)= .D DI D Do u (t) is the function of the output signal of PWM module, q(kT) is the Kth value of the input signal of PWM module, and D is a duty, which is determined as I q(kT)/qmaxl. Do is the duty value corresponds to the delay time in the valve operation. A vibration which continues steadily with a constant amplitude and frequency is called a limit cycle. Especially, when the ratio of (Tq/ T) is 2 as shown in Fig. 3, we call it 1 pulse IT I (k -G. Fig. 3 ? i l)T (k+2)T Output pulse trains of 1 pulse limit cycle. fi -0 (k+n)r (k+n+1)T (k+2n-1)T kT (k+l)T (k+n-1)T t Fig. 4 Output pulse trains of n pulse limit cycle. limit cycle. The Fourier coefficient al and bl of basic vibration is given by 2s sin( r (c7) (11) al- 7/ 2s / r 1 b, =-LCOST. r () - (12) then the describing function of PWM signal gen- eration is given as the following for the case of 1 pulse limit cycle. r 2s . ;r , NwM(#, l, TJ-Tgksn,r ) ) Meanwhile, the n pulse limit cycle has succes- sively n positive pulses and n negative pulses for one period time(Tq) of limit cycle as shown in Fig. 4. The describing function of this case is derived as the following, 2s . n-t . NpwM (c7, n, b)= _ e-k_oe- lrq = - e-Jk+ nT r(q(kT) (14) As shown in Eq. (14), NPwM(ff, n, b) is a function of the amplitude () of the input signal of PWM module, the pulse number(n) of limit cycle and synchronizing degree(b), but it is in- dependent to angular frequency (w). 2.3 Describing function of friction charac- teristic The friction characteristic of hydraulic cylinder which is modelled as shown in Fig. 5 consists of static friction, coulombs friction and viscous damping friction which is dependent upon piston velocity. Supposing the hysterisis in friction char- acteristic is negligible, the friction characteristic shown in Fig. 5 can be formulated as the follow- ing: (Back6, 1992) I ( Fe()=Gk. + F+L. l- sgn(,7) (15) A Study on the Stability Analysis of a PWM Controlled Hydraulic Equipment 401 In order to derive the describing function, the above equation could be approximated as (Proch- nio, 1986) FR():Gg+Fc4 L 5 , sgn(v) (16) Where c is a constant required to approximate Eq. (15) to Eq. (16). In harmonic balance, the piston velocity v(t) is given by v(t) = 7-sin(wt) (17) the describin function of friction characteristic is NR() = a(g) (18) qvhere a(7) is the Fourier coefficient of basic vibration as follows: 1 ,-2r _ +F a(g)=-J 1Gv sin(x) L Ls + 1 +( g sin(X)c .)2- sgn(sin X) sin(x) y Fig. 5 Model of cylinder friction. Eq. (18) the describing function of friction char- acteristic becomes as follows: NR () =Gk+ 4Fc 4 2Lsc 2 . In xv x2/ 2 + c2 t7 + (20) 2.4 Analysis of the control loop In order to investigate the stability of the system which have highly nonlinear characteristics such as PWM signal generation and friction in cylin- der, the describing functions of these nonlinear characteristics are derived. Hence, the well known linear analysis method can be utilized. The above derived Eqs. (6), (7), (8), (14) and (20) are combined to construct the block diagram of the overall system as shown in Fig. 6. In this figure the nonlinear characteristics which is depicted by the describing functions are shown with double square frame. The transfer function Ns of the valve-cylinder system is derived as Eq. (21), the input of which is the valve spool displacement X and the output is cylinder displacement x, Ns(W, g)=- c (21) s ( s2 + c2s + c3 where, C1- k c2= FRM c _ 4J - VM The unstable phenomenon of a vibration sys- tem which has an element with a highly nonlinear characteristic tlsually starts with the limit cycle, Nsr ), F Fig. 6 Block diagram of PWM hydraulic cylinder system. 402 Jun-Young Huh and G. Wennmacher that is, the output of the system oscillate continu- ously with a constant amplitude and frequency. Whether the limit cycle will occur or not in a certain condition can be determined through the investigation of the solution of the closed loop characteristic equation. That is, in the harmonic vibration balance state, the limit cycle will occur if the solution of the closed loop characteristic equation exists (Prochnio, 1986; Follinger, 1991). The objective of this study is to present a method to predict the stability of the system so that it can be used to guarantee the stable operation in a wide range of operating conditions. The charac- teristic equation of the position control system shown in Fig. 6 is given as 1 + Ke - 9 NewM Ns :0 (22) where K is obtained by the product of the propor- tional gain(k1) and the maximum displacement of valve poppet(kin). The characteristic Eq. (22) is so complicated that it is very difficult to solve it analytically. Therefore, a graphical method is utilized to get the solution of the characteristic equation. wave(T) comes to 2n (n=l, 2, 3, .-.), hence, the relation between angular frequency(co) and PWM carrier wave period(T) is derived as the following: z n=l, 2, 3, . (24) co= nZ In the above equation, when the pulse number (n) of limit cycle is fixed, angular frequency(co) is uniquely determined. Therefore the unfixed parameters of Eq. (22) are the gain(K), the amplitude () of the input signal of PWM mod- ule, synchronizing degree(e) and pulse number (n) of limit cycle. At first to investigate whether the 1 pulse limit cycle occur or not, the pulse number of limit cycle is set as n=l. Then, we changed K value variously in the range of interest for the fixed r and value, otherwise we chan- ged value variously for the fixed r and K value, and investigated the solutions which satisfy the Eq. (22). And for the cases of n=2, 3, . it is investigated in the same way. The values of each parameters which is used in computer simu- lation are shown in Table 1. 3. Computer Simulation and Experiment In order to solve the characteristic Eq. (22) by graphical method which is utilized in this study, the first step is to investigate the Eq. (22) and look for the unfixed parameters. In the second term of the left hand side of the Eq. (22) NewM (c7, n, r is the function of the amplitude(p) of the input signal of PWM module, the pulse num- ber(n) of limit cycle and synchronizing degree (r but it is independent to angular frequency (co). Ns(co, g) is the function of angular fre- quency(w) and piston velocity amplitude(f). The relation between the amplitude(p) of the input signal of PWM module and piston velocity amplitude is given by t = k/ g (23) co When this system is operated in harmonic vibra- tion balance state, the ratio between the input signal period (Tq) and the period of PWM carrier Table 1 System parameters used in computer simulation. Parameters Value Dimension A 7.65 cm 2 c 1.5 cm/s Fc 0 kgf L 21 kgf G, 5.46 kgf s/cm kl 4666.7 cm2/s k2 0. 3536 cmS/kgf s k, 0.015 cm M 5-80 kg T 5 ms Ts 1 ms Vt 306 cm 3 Va 50 cm 3 r 1 ms fie 12000 kgf/cm 2 A Study on the Stability Analysis of a PWM Controlled Hydraulic Equipment 403 I I I I I I I- ! J I r-1 I I I ! I I I I I I I I I I I Ace.2 I I I I I I I I L. -3 r I I I I I I .3 Fig. 7 Hydraulic circuit of experimental equipment. The experimental apparatus is constructed as shown in Fig. 7, which has 4 high speed on-off valves, a double rod hydraulic cylinder, and a microcomputer as the major components. Piston displacement is measured by the digital optical Table 2 Specification of ex 9erimental apparatus Equipment Specification Hydraulic System Electric Equipment Electric Motor 5 PS Hydraulic Pump Qmx:8.9 l/min Actuator Ak = 7.65 cm 2 St.= 10 cm High speed On-off valve Relief valve Filter Accmulator 1 Accumulator 2 Displacement Transducer Microcomputer Controller Qmax:4.2 1/min Xvmax:0.015 cm ton,dead :0.46 ms /on,dlp :0.18 ms toll,dead = 0.30 ms tou,dlsp :0.16 ms Pmax :230 kgf/cm 2 10 m 4 liter 0.075 1, 70 bar 12 bit (0.315ram) CPU386 sensor with + l#m accuracy. Pressure transducer is represented by P/T in Fig. 7, of which accuracy is - bar. It is an analog sensor so that the analog to digital converter is required to transfer measured values in the computer. Tile controller in control loop gets the sampled data on every millisecond, calculates the control law and gives it to 4 channel PWM modulator. PWM carrier frequency is 200 Hz and the period is 5ms. For the various inertia loads, we looked for the propor- tional gain that makes the system on the threshold from stability to instability with limit cycle. The major specifications of the components which is used in this experiment are shown in Table 2. 4. Performance of the Control System In the cases of inertia load 6kg and 50kg, the experimental results for step input are showed with the solid line and the dotted line respectively in Fig. 8 (ks= 1 5). The responses shows instabil- ity with almost regular oscillations which have almost constant amplitudes. For the inertia load 6kg, we can find that the response shows 1 pulse limit cycle because it vibrated with 100Hz fre- quency which have 10 peaks per 0.1 second, and 404 Jun-Young Huh and G. Wennmacher ? i. ,. . I . . . i . i i ; i i . 0.5 ; i i ! 0 0 0.05 0.1 0.15 0.2 .me (sac) Fig. 8 Experimental results of n pulse limit cycle (n= 1, 2). this is the half of the PWM carrier frequency 200Hz. For the inertia load 50kg, the response shows almost 2 pulse limit cycle to have about 5 peaks per 0.1 second. Here the reason that the response does not show the exact 5 peaks per 0.1 second is guessed due to the followirtg vibration of high speed on-off valve after its opening and closing operation. Figure 9 is one of the result figures which is utilized to investigate analytically the existence of 1 pulse limit cycle which is the solution of Eq. (22). One solid line of Fig. 9 (a) is obtained for the case with the inertia load of 11 kg, synchronizing degree (b) of 88 and propor- tional gain(ks) of 0.7, by plotting the value of Ke -s 9 NpWM Ns in s-plane for the variation of the velocity amplitude(7) from 0.01 to 25, and the other solid line curves are obtained by repeat- ing this process for the increased proportional gain(ks) from 0.7 to 1.5 by step of 0.2.Similarly when the proportional gain (kl) varies from 0.1 to 4.5 for the increased amplitude(O) from 0.2 to 3.2 by the step of 0.6, the values ofKe - 9 NpwM 9 Ns are plotted with dotted lines. Herein, when the curve crosses the point (-1, 0) of s-plan, there exist the solution of Eq. (22) which is given by the values of the proportional gain (kj) and veloc- ity amplitude (7). One solid line of Fig. 9 (b) is obt
收藏