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機(jī)電工程學(xué)院
畢業(yè)設(shè)計(jì)外文資料翻譯
設(shè)計(jì)題目: ZY1160貨車底盤總體及車架設(shè)計(jì)
譯文題目: 汽車工程學(xué)I：汽車縱向動力學(xué)
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正文：（選自《汽車工程學(xué)Ⅰ：汽車縱向動力學(xué)》P111--122 ）
變速器
3.4.1.2 變速器設(shè)計(jì)
機(jī)械有級式變速器被分成兩類：
1）雙軸式變速器
2）周轉(zhuǎn)式變速器（行星齒輪變速器）
雙軸式變速器又被稱為中間軸變速器，這是適用于所有傳統(tǒng)有級式變速器的通用術(shù)語。在這種變速器中。在這種變速器中，非直齒輪之間的轉(zhuǎn)矩傳遞發(fā)生在兩個(gè)平行軸上相互嚙合的齒輪對。
前進(jìn)擋的直齒輪對是常嚙合的。這個(gè)齒輪對是由一個(gè)與軸固定在一起齒輪與一個(gè)空套在軸上的齒輪組成。后者是在第二軸上并且可以自由旋轉(zhuǎn)但是不能軸向移動。同步器可以按照各個(gè)不同要求使空轉(zhuǎn)的齒輪對與軸相連進(jìn)行傳動。
雙軸式變速器可以根據(jù)輸入和輸出軸的安裝位置被分為：
1）共軸雙軸變速器
2）非共軸雙軸變速器
（1）共軸雙軸變速器 對于傳統(tǒng)驅(qū)動形式（發(fā)動機(jī)前置后驅(qū)）的車輛，這種雙軸有級式變速器被廣泛的采用。這種設(shè)計(jì)的顯著特點(diǎn)是可以把輸出軸與輸入軸直接相連進(jìn)行傳動。剩余的能量齒輪對間的能量傳遞是通過中間軸進(jìn)行的（圖3-52）。
圖3-52 共軸式雙軸變速器的能量傳遞路線及其布置
（來源：電機(jī)專用驅(qū)動系統(tǒng)）
一般來言，共軸的中間軸變速器直接與離合器總成裝配在一起。輸出轉(zhuǎn)矩通過傳動軸進(jìn)行傳遞。這將導(dǎo)致不合理的軸荷分配。正是這樣的原因在有些車輛上將變速器安裝在后軸附近（驅(qū)動橋設(shè)計(jì)).傳動軸的轉(zhuǎn)速與發(fā)動機(jī)發(fā)動機(jī)的轉(zhuǎn)速一致。
圖3-53展示了一種客車上使用用的共軸的五擋變速器。
圖3-53 五擋手動變速器（福特）
圖3-54 非共軸式雙軸變速器的能量傳遞路線及其布置
（ 來源：電機(jī)專用驅(qū)動系統(tǒng) ）
圖3-55 非共軸五擋手動變速器（薩博）
（2）非共軸的雙軸變速器 這種雙軸變速器被用于發(fā)動機(jī)前置前驅(qū)或者發(fā)動機(jī)后置后驅(qū)的車輛，這種變速器的輸入輸出軸常常被安置在變速器的同一邊?？拷敵鲚S的是差速器，差速器是集成在變速器殼上的。這種變速器的動力傳遞路線如圖3-54所示。每個(gè)擋位中，動力傳遞是通過一對嚙合的齒輪對進(jìn)行的。這種變速器對于前輪驅(qū)動的車輛來言是不合適的因?yàn)榍S與輸入軸之間的軸向移動必須被橋接。
圖3-55展示前輪驅(qū)動且發(fā)動機(jī)橫置的汽車的非共軸式雙軸變速器的實(shí)例。
3）換擋機(jī)構(gòu) 在雙軸變速器中是通過適當(dāng)?shù)碾x合機(jī)構(gòu)來進(jìn)行換擋的。在這種情況下，有一些不同：
1) 同步器
2）粘附換擋離合器
在手動變速器中換擋是通過換擋桿進(jìn)行操作的。在換擋的時(shí)候，換擋齒輪和軸的同步時(shí)間是確定的，齒輪和軸的速度必須是相同的。這種無同步器換擋的方法曾經(jīng)被用了很長時(shí)間，這是通過駕駛員兩次的離合操作來實(shí)現(xiàn)速度的同步的。對于無同步器的變速器來言選擇正確的換擋時(shí)機(jī)是需要一定技巧的。這就導(dǎo)致了現(xiàn)在這種換擋方式僅僅被應(yīng)用在一些卡車上。在有級式變速器的前進(jìn)擋中是通過摩擦部件來保證齒輪軸向移動過程中旋轉(zhuǎn)部件速度的同步來實(shí)現(xiàn)操作的方便和快捷以及低噪聲和少磨損的目的的。
圖3-56 博格華納系統(tǒng)單錐同步器（ZF）
1---在滾針軸承支持下空轉(zhuǎn)的惰齒輪 2---有摩擦錐面的結(jié)合齒圈
3---主要功能元件，有反錐面的摩擦面及起鎖止作用的齒的鎖環(huán)。
4---能夠使變速器軸與結(jié)合套相互聯(lián)系一起轉(zhuǎn)動的花鍵轂
5---壓緊彈簧 6---球銷 7---推力塊 8---具有內(nèi)齒的結(jié)合套
圖3-57展示了鎖環(huán)式同步器的工作原理。同步器的主體花鍵轂4與變速器的輸出軸相聯(lián)。鎖環(huán)3被花鍵轂帶動一起旋轉(zhuǎn)。鎖環(huán)上有比花鍵轂上槽窄的凸起，可以保證鎖環(huán)可以相對于花鍵轂有一定的轉(zhuǎn)動角度。
在換擋開始之前，撥叉被止動器保持在中間位置。換擋時(shí)，撥叉強(qiáng)制移動結(jié)合套8，這會使滑塊7帶動球銷6向鎖環(huán)3施加壓力。鎖環(huán)的錐面與齒圈上的錐面相抵觸。與齒輪1有關(guān)的結(jié)合套8與鎖環(huán)3之間的速度差會使齒輪1與鎖環(huán)的速度相等，這就是同步過程。
換擋撥叉進(jìn)一步移動，結(jié)合套8上的斜面與鎖環(huán)3上對應(yīng)齒端倒角斜面相接觸。重要的同步動作開始，就像圖2展示的那樣，換擋力通過滑塊7與凸牙8作用于同步環(huán)上。換擋力被分解成兩個(gè)力產(chǎn)生在斜面之間的拔環(huán)力矩TZ有使鎖環(huán)后退的趨勢。在存在摩擦的過程中不能進(jìn)行換擋。經(jīng)計(jì)算可得當(dāng)拔環(huán)力矩TZ小于作用在鎖緊裝置上的摩擦力矩Tr的階段將會是鎖止裝置鎖死。在文獻(xiàn)中常常稱拔環(huán)力矩TZ為轉(zhuǎn)位轉(zhuǎn)矩Ti，稱摩擦力矩Tr為錐面力矩Tc。
當(dāng)速度同步的時(shí)候摩擦力矩為零，如圖3 。解鎖過程開始，拔環(huán)力矩大于摩擦力矩，使鎖環(huán)相對于齒輪錐面轉(zhuǎn)動一個(gè)角度。在這個(gè)階段，換擋力迅速的減少。通過換擋撥叉的軸向移動彈簧支承球銷進(jìn)入結(jié)合套的凹槽中，這把彈簧5按壓到滑塊中，直到被結(jié)合套覆蓋。
圖3-57 同步過程。用半滿的提示箭頭指示的方向移動，扭矩箭頭指示作用同步器環(huán)上的力矩
當(dāng)齒輪嚙合時(shí)，結(jié)合套上齒越過鎖環(huán)使花鍵轂2與齒圈嚙合。在圖4中球銷被覆蓋。鎖環(huán)只有通過推力塊施加于花鍵轂進(jìn)而施加在摩擦錐面上的殘余壓力。殘余壓力產(chǎn)生于移動結(jié)合套和滑塊(包括球銷）移動時(shí)產(chǎn)生的摩擦力。結(jié)合套相對與鎖環(huán)轉(zhuǎn)動一個(gè)角度而與花鍵轂結(jié)合。換擋動作完成。結(jié)合套把轉(zhuǎn)矩由齒輪傳到變速器軸上，如圖5.
圖5-38五擋雙離合變速器（保時(shí)捷）
同步的過程無論如何都是要耗費(fèi)一些時(shí)間的。為了縮短換擋的時(shí)間，保時(shí)捷公司已經(jīng)研發(fā)能過在受載時(shí)還能實(shí)現(xiàn)換擋的雙離合式變速器很長時(shí)間了。在以一個(gè)擋位前進(jìn)時(shí)，相鄰的高檔位或者低檔位已經(jīng)被換上，但是真正的換擋發(fā)生在分離當(dāng)前離合器而同時(shí)結(jié)合另一個(gè)離合器。這意味著這這種換擋操作是基于摩擦的換擋操作。圖3-58展示了非共軸雙離合器式變速器布置圖。
第二個(gè)離合器通過空心軸與齒輪2和4相聯(lián)，第一個(gè)離合器通過空心軸的內(nèi)軸與齒輪1,3和5相聯(lián)。因此，兩個(gè)連續(xù)的擋位可以被同時(shí)以合適的嚙合時(shí)機(jī)被換上。但是無論如何在常規(guī)的雙軸變速器中在不中斷動力傳遞是進(jìn)行換擋是不可能實(shí)現(xiàn)的。
為了有足夠的換擋速度和舒適性，離合器的電動自動控制換擋機(jī)構(gòu)是必不可少的。此外，駕駛者還可以直接手動操作。
2. 周轉(zhuǎn)式變速器（行星齒輪變速器）
周轉(zhuǎn)式變速器的特點(diǎn)是至少三個(gè)同軸桿通過各自齒輪的嚙合而被永久的聯(lián)系在一起。圖3-59展示一個(gè)周轉(zhuǎn)式變速器上不同中心軸上齒輪的各自位置。
兩根軸分別與太陽輪和齒圈相聯(lián)，第三根軸與行星架相聯(lián)。為了實(shí)現(xiàn)較小的軸向尺寸，采用簡單的齒輪形式以及較好的受力情況而選擇了這種設(shè)計(jì)。這種周轉(zhuǎn)式變速器設(shè)計(jì)被稱為直齒行星齒輪變速器。
圖3-59 行星齒輪變速器
周轉(zhuǎn)式變速器的特征參數(shù)是公稱變速比i0，它是齒圈與太陽輪的半徑比。
周轉(zhuǎn)式變速器的傳動比可由下面的基本方程得到：
這里 -----行星架的角速度；
-----齒圈的角速度；
------太陽輪的加速度；
------公稱變速比。
圖3-60展示了在穩(wěn)定情況下行星齒輪的受力情況。
圖3-60 作用在行星齒輪上的力
由行星齒輪上的力和力矩平衡可以得到：
這里
考慮到各個(gè)量的大小，我們可以得到：
因?yàn)?
我們可以類似的得到剩余的傳動比：
因此，如果忽略傳動過程中的能量損失，三根軸的轉(zhuǎn)矩比可以僅僅用行星齒輪機(jī)構(gòu)的幾何參數(shù)來確定。
因?yàn)榱仄胶?我們可以得到
圖3-61 行星齒輪變速比以及理論變速比的變化范圍
因?yàn)檗D(zhuǎn)矩即不會增加也不會減少，這樣，行星架轉(zhuǎn)矩MB的方向與太陽輪轉(zhuǎn)矩MS方向以及齒圈轉(zhuǎn)矩MH方向的方向相反。當(dāng)所有軸都向一個(gè)方向轉(zhuǎn)動時(shí)行星齒輪傳動的輸出也同樣適應(yīng)這種情況。
與傳統(tǒng)有級式變速器相比，自由轉(zhuǎn)動的行星齒輪變速器不受力。如果一個(gè)軸被固定，它成為力矩傳遞的支承。另兩根軸之間的轉(zhuǎn)速比固定這個(gè)比值可由它的特征方程決定。依靠與發(fā)動機(jī)相聯(lián)的軸以及被固定的轉(zhuǎn)速器殼，行星齒輪傳動機(jī)構(gòu)可以有六個(gè)不同的傳動比，其中兩個(gè)是倒擋。
而且，還可以通過第三根軸直接與發(fā)動機(jī)相聯(lián)而獲得直接擋。所有的行星輪作為一個(gè)整體旋轉(zhuǎn)。傳動比i=2,i=0.5以及i=-1不能通過簡單的圓柱行星齒輪來實(shí)現(xiàn)。為了實(shí)現(xiàn)一定的傳動比，齒圈的齒數(shù)應(yīng)該根據(jù)太陽輪的齒數(shù)來選定。這可以用錐齒輪或者準(zhǔn)配多排行星齒輪變速機(jī)構(gòu)。
圖3-62 行星齒輪變速器上制動齒圈的多片離合裝置
由于所受作用力被分布到幾個(gè)嚙合的齒輪（幾個(gè)行星輪）上，使這種優(yōu)良的對稱設(shè)計(jì)同時(shí)具有了較小的體積以及較輕的質(zhì)量和較大的最大傳遞轉(zhuǎn)矩。因此大量的行星齒輪變速器被應(yīng)用在汽車上（自動變速箱，分動器，輪邊減速器）。
在這種設(shè)計(jì)中包括一些可變的傳動比，行星齒輪變速器可以用制動離合機(jī)構(gòu)來實(shí)現(xiàn)在承載的條件下?lián)Q擋。在現(xiàn)在的變速器中，這些換擋單元有多盤制動器和離合器以及帶式制動器組成（圖3-62）它們大多是由液壓系統(tǒng)操縱的。
為了使變速箱中有較少的換擋機(jī)構(gòu)并且有較多的擋位以及能夠很好地設(shè)計(jì)各擋位之間傳動比的關(guān)系，雙排行星齒輪變速器出現(xiàn)了。
圖3-63 雙排行星齒輪變速器的設(shè)計(jì)和擋位安排（戴姆勒奔馳）
C--- 離合器 B---制動器
圖3-63展示了4擋雙排行星齒輪變速器的原理示意圖以及其適合的變速離合機(jī)構(gòu)。
通過一個(gè)太陽輪輸入，在后排行星傳動機(jī)構(gòu)鎖止的時(shí)候通過兩個(gè)太陽輪輸入，通過后排行星齒輪傳動機(jī)構(gòu)的行星架輸出。傳動比根據(jù)行星齒輪傳動機(jī)構(gòu)的數(shù)量多次重復(fù)應(yīng)用特征方程求得。
在自動變速器中，采用的雙排行星齒輪變速器。圖3-64展示了一個(gè)各擋位傳動比排列合理并且占用空間較小的拉維挪式變速器，它的設(shè)計(jì)原理及換擋機(jī)構(gòu)簡圖在圖3-65上展示。
圖3-64 拉維挪式變速器 （來源：電機(jī)專用驅(qū)動系統(tǒng) ）
由太陽輪輸入由齒圈輸出就像圖3-64展示的那樣。短的和長的行星輪用一個(gè)行星架連接。
圖3-65 拉維挪式變速器的設(shè)計(jì)及擋位安排（博格華納）
在有一個(gè)固定傳動比的圓柱行星齒輪傳動機(jī)構(gòu)被用作輪邊減速器。這樣輸入軸的轉(zhuǎn)矩低于輸出軸的轉(zhuǎn)矩，且轉(zhuǎn)速高于輸出軸轉(zhuǎn)速。減速器被安裝在車輪上會減少非簧載質(zhì)量（對比：車輛工程Ⅱ）。用雙級主減速器，可以明顯減少“短”半軸的長度就像一些工程車輛，而且因?yàn)樾枰^小傳動比的主減速器可以使空間的利用率提高。一個(gè)缺點(diǎn)就是由于多級傳動不可避免地減少了傳動效率，所以輪邊減速僅僅被用于商用車（圖3-66）。
行星齒輪變速器也用于分動器（3.5節(jié)）或者增距變速器（在功率分流設(shè)計(jì)中）周轉(zhuǎn)式變速器中沒有中間軸被固定。在分動器中輸入轉(zhuǎn)矩被分成兩個(gè)輸出在增距變速器中輸入轉(zhuǎn)矩可以被相加之后進(jìn)行輸出，可是在這種情況下，周轉(zhuǎn)式變速器不再是一個(gè)功率轉(zhuǎn)換器，這是因?yàn)樯倭酥С辛亍?
圖3-66 行星齒輪輪邊減速器
3.4.1.2 Transmission designs
Mechanical stepped transmissions are classified into two design:
1) Dual shaft transmission
2） Bicycle transmission
1.Dual shaft transmission
The dual shaft transmission , also called the counter shaft transmission ,is the generic term applicable to all conventional stepped transmissions . Here , the torque transfer in the non-direct gear level takes place over externally toothed spur gear pairs mounted on two parallel shafts .
The spur gear pairs for the forward gear permanently mesh with each other . They consist Ufa fixed gear wheel connected to a shaft and a free gear wheel . The latter is positioned on the second shaft and can freely rotate but cannot be axially displaced . Positive-engaging clutch elements are used to engage the freely rotating gears to the shaft , as per the demand .
Dual shaft transmissions can be classified according to the position of the transmission input and output shaft into :
1) Coaxial two-shaft transmission
2）Deaxial two-shaft transmission
(1) Coaxial dual-shaft transmission In case of standard-driven vehicles (front engine with rear wheel drive ) , the dual-shaft mechanical stepped transmission in coaxial design is predominatly used . The coaxial position of the input and the output shaft which can normally be connected rigidly in direct gear is a characteristic feature of this design . The power flow for the remaining gears is led over a counter shaft (Fig.3-52) .
Fig.3-52 design and power flow in a coaxial dual-shaft transmission
( Source : mot - Special , Anterior )
As a rule , the coaxial counter shaft transmission is directly coupled to the engine along with the clutch assembly . The output is led over the cardan shaft . This leads to an unfavorable axle load distribution. This is the reason for the transmission being located at the real axle in some vehicles ( transaxle design ) . The cardan shaft then rotates at the engine speed .
Fig.3-53 shows an example for a coaxial five-speed transmission of a passenger car .
Fig.3-53 Five-speed manual transmission ( Ford )
(2) Deaxial dual-shaft transmission In dual-shaft transmission ,for vehicles with front-wheel drive or rear wheel driven vehicles with rear engines , the transmission input and output shaft have an axial offset and are normally arranged on the same side of the transmission . Close to the output is the differential , which is integrated into the transmission housing . The power flow in dual-shaft transmissions , in deaxial design , is indicated in the functional diagram shown in Fig.3-54 . In each gear , power is transferred over a gear wheel pair . Such a transmission is advantageous for front wheel driven vehicle since an axial offset between the crankshaft position and the position of the drive shaft must always be bridged .
Fig.3-54 Design and power flow of a deaxial two-shaft transmission
( Source : mot-Spezial , Antried )
Fig.3-55 shows an example of a deaxial dual-shaft transmission with front wheel drive and a transversely-mounted engine .
(3) Gearshift In dual-shaft transmission , shifting into different gears takes place by means of appropriate shifting clutch agent . In this case , one differentiates between :
1) Positively-engaged shifting clutch agents .
2）Adherent shifting clutch agents .
Fig.3-55 Deaxial 5-speed-manual transmission ( Saab )
In manual transmission , shifting of gears takes place over the gearshift roads . Using displacing sleeves , a positively-engaged link between the gear wheel and the shaft is established . A speed synchronisation of the shaft and the gear wheel is required . The so called non-synchromesh transmissions were used for a long time , where the speed equality was achieved by double declutching by the driver . The correct handling of non-synchromesh was achieved requires a certain skill , the result being that today ,they are only occasionally used in trucks . In order to achieve convenient operation at high shift speeds ,low noise and low wear , the forward geard gears of stepped transmissions are synchronised by ensuring speed equality of the rotating parts which have to be engaged during gear shift , over the frictional elements . Fig.3-56 show the elements involved in speed synchronisation .
Fig.3-56 borg-warner system single-taper synchroniser ( ZF )
1---idler gear running on needle bearings
2---synchroniser hub with selector teeth and friction taper
3---main functional element, synchroniser ring with counter-taper and locking toothing
4---synchroniser body with internal toothing for positive locking with the transmission shaft and external toothing for the gear shift sleeve 5---compression spring 6---ball pin
7---thrust piece 8---gearshift sleeve with constant-mesh internal gearing
Fig.3-57 shows the functional principle of the locking synchronisation . The synchroniser body 4 is fixed to the transmission shaft . The synchroniser ring 3 is guided by stop bosses in the syschroniser body . These are narrower than the grooves in the synchroniser body , which allows the synchroniser ring a certain amount freedom to twist radially .
Before the shifting process starts , the gearshift sleeve is held in the middle position by a detent . The gearshift force F triggers the axial movement of the gearshift sleeve 8 , which causes the thrust pieces 7 to act on the ball pin 6 to press the synchroniser ring 3 with is counter-taper against the friction taper of the synchroniser hub 2 , the speed difference between the gearshift sleeve 8 and the synchroniser ring 3 relative to the idler gear 1 causes the synchronising process which is known as “ asynchronsising “ ( Fig.3-57 ) .
Fig.3-57 Synchronising process . Arrows with half-filled tips indicate the direction of movement , the torque arrows indicate the torques acting on the synchroniser ring
The gearshift sleeve is moved further . This brings the bevels of the constant-mesh internal gearing of the gearshift sleeve 8 and the constant-mesh external gearing of the synchroniser ring 3 into contact . The main synchronisation action starts , phase 2 . The gearshift force is applied to the synchroniser ring via the thrust pieces 7 and the dogs 8 , the force being divided between them . The gearing torque TZ arising at the bevels acts to close the locking device . In the slipping phase the gearshift sleeve cannot be shifted . In the literature the gearing torque TZ that acts to close the locking device . In the slipping phase the gearshift sleeve cannot be shifted . In the literature the gearing torque TZ is frequently referred to and the index torque Ti and the friction torque TZ as the taper torque Tc . When speed sysnchronnisation has been achieved , the gearing torque becomes greater than the friction torque , and acts via the bevels to return the synchroniser . The gearshift force decrease rapidly in this phase . Throughout the axial movement of the gearshift sleeve , the spring-loaded ball pin slides along the inclined grooved surface . This presses it against the spring 5 into the thrust piece , until it is covered by the gearshift sleeve .
As the gear is engaged , the gearshift sleeve toothing encounters the bevels of the selector teeth of the synchroniser hub 2 . In this phase 4 the ball pin is covered . The synchroniser ring is pressed against the friction taper of the synchroniser hub only by residual pressure via the thrust pieces . This residual pressure arises from the friction between the moving the moving gearshift sleeve and the trust pieces ( with ball pins ) . The gearshift sleeve toothing twists the synchroniser hub relative to the synchroniser ring . The shift movement is enable . The gearshift sleeve positively engages the power flow between the gear pair and the transmission shaft , phase 5 .
This process of synchronisation however requires finite time . In order to shorten this synchronisation shifting time , Porsche worked for many years on the so called “ double clutch transmission “ which can be shifted under load . Driving in a particular gear , the next higher or lower gear can be shifted . The real gearshift takes place by opening one clutch and closing the other simultaneously . This means that the gearshift in this case is friction based . Fig.3-58 shows the layout of a double clutch transmission in deaxial design .
The second clutch is connected over a hollow shaft with the gear wheels of the gears 2 and 4 and the first clutch over the inner shaft with the gear wheels of the gears 1 , 3,and 5 . Thus , two successive gears can be shift between the gears on one shaft without torque interruption , as is the case in regular dual-shaft transmissions .
For sufficient speed and comfort , electronic regulation of clutch control and gearshift is necessary . The driver , additionally , can externally influence this arrangement .
Fig.3-58 5-speed double clutch transmission ( Porsche )
(2) Epicyclic transmission ( planetary gearbox )
Epicyclic transmissions are characterised by at least three coaxial center shafts that are permanently engaged each other over the gear wheel . Fig.3-59 shows the different gear wheel positions in an epicyclic transmission with three center shafts .
Two of the shafts involved are connected to the central wheels , the sun wheel and the hollow wheel . The third is coupled to a bar which is connected to the so-called planetary wheels
which mesh with both the central wheels . This design is chosen , when possible , as a result of its low constructional length simplicity of the gear wheels and the point of force application which are in place . This design of the spur-gear planetary transmission .
Fig.3-59 Epicyclic transmission
The characteristic parameter of an epicyclic transmission is the nominal ratio i0 which is defined as the ratio of hollow wheel radius to the sun wheel radius .
The speed ratios of a planetary gear train result from the basic equation for epicyclic transmission :
Where -----angular velocity of the bar ；
-----angular velocity of the hollow wheel ；
------angular velocity of the sun w ；
------nominal ratio 。
Fig.3-60 shows the forces on a planetary wheel i.e. In a constant state of motion .
Fig.3-60 Forces acting on a planetary wheel
The balance of forces and torques on a planetary wheel results in :
Where
Considering the magnitude , we get :
Since
The remaining torque ratios can be calculated similarly :
Thus , when losses during moment transfer are neglected , the ratio of torques on the three center shaft can be clearly determined , using only the geometric data of the planetary train .
For equilibrium , we obtain
Hence the torques are either split up or added . In case , the direction of bar torque MB is opposite to the torques MS and MH acting in the same direction . When all shaft rotate in the same direction , the same is applicable to the power output of a planetary transmission .
In contrast to the usual stepped transmissions , free bearing forces . If a shaft is rigidly braked ,it acts as a torque support . The other two shafts move relative to each other in a firm speed ratio , which can be determined from the basic equation . Depending on the coupling of the individual center shaft with the engine , the output and the rigid transmission housing , a planetary train allows for six different ratios , two of which reverse the direction between the engine and the output sides ( Fig.3-61 ) .
Fig.3-61 steps of a planetary gear with theoretical range ratios
Moreover , a direct gear can be realised if the third shaft is linked with the engine or the output . The entire planetary gear then rotates as a block . The ratios i=2 , i=0.5 or i=-1 cannot be realised by a simple spur-gear planetary . In order to achieve this , the number of teeth on the hollow gear have to correspond to that on the sun gear . The situation can be set right using conical gear wheels or by assembled epicyclic transmission .
The favorable , symmetrical design as well as the low constructional volume and low weight with a simultaneously high maximum transferable torque , due to force distribution over several meshings ( several planetary wheels ) , led to a large number of applications of planetary transmissions in motor vehicles ( automatic transmission , transfer case in all-wheel drive systems , wheel-hub transmission ) .
In designs which involve several ratios , the planetary transmission can be shifted under load by employing adherent clutch shift agents . In realised transmissions , these shift elements consist of multi-disk brakes and clutches as well as band brakes ( Fig-3-62 ) , which for the most part are hydraulically operated .
In order to have less number of shift clutch agents in gearboxes with high number of gear ratios and to freely design in the relationship of the individual gear ratios against each other , coupled planetary transmissions used .
Fig-3-63 shows the principle sketch of a 4-speed transmission with 2 planetary sets and the appropriate shift clutch agents .
Fig.3-62 Adherent shift clutch agents in epicyclic transmission for braking the hollow wheel
Fig.3-63 Design and shifting diagram of a coupled planetary train ( Daimler Benz )
C-----clutch b-----brake
The input is performed over one sun gear , or in case of a blocked second planetary train , over two sun wheels . The output takes place over the planetary carrier of the rear planetary train . The particular gear ratios result from from the repeated use of the basic equation , according to number of planetary trains involved .
Fig.3-64 Ravigneaux-transmission ( Source : mot-Spezial , Antrieb )
In automatic transmission , often special designs of coupled planetary transmission are used . Fig.3-64 shows a Ravigneaux-transmission that has a simple design of favorable gear ratios and a low constructional space . The design and shifting diagrams are shown in Fig-3-65 .
Fig.3-65 Design and shifting diagram of Rvigneaux-transmission ( Borg Warner )
C---clutch B---brake O---one-way clutch
The input is performed over the sun wheels , while the output is derived over the hollow gear in the transmission as shown in Fig.3-64 . The short and long planetary gears are are connected by a common bar .
In non-shiftable designs with a fixed gear ratio , spur-gear planetary trains are used as wheel hub transmissions .