JP2005023987A - Multistage shift planetary gear train - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve the applicability to the transmission of a front-wheel-drive vehicle with a horizontal engine by allowing a reduction in the axial length of a gear train while using two sets of planetary gears for producing a shift ratio with forwarding six or eight stages. <P>SOLUTION: The multistage shift planetary gear train comprises a first planetary gear set 14 on an input shaft 10 and a second planetary gear set 18 on a driving shaft 16. The input shaft 10 and the driving shaft 16 are connected to each other via a pair of first gears 40, and a first sun gear 20 and a second sun gear 30 are connected to each other via a pair of second gears 58 and can be fixed to each other. The second sun gear 30 can be connected to the driving shaft 16. A first ring gear 22 and a second carrier 38 are connected to each other via a pair of third gears 60 and connected to an output shaft 12. A first carrier 28 can be fixed and connected to the input shaft 10, and a second ring gear 32 can be connected to the driving shaft 16. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

前進６段ではあるが、車両用変速機にふさわしい変速比が得られ、しかも従来例より軸方向長さを短くできることは言うまでもない。
【００５０】
次に、本発明の多段変速遊星歯車列における第２の実施の形態のスケルトンを図４に示す。
ここでは、図１に示した実施の形態と異なる部分を中心に説明し、図１の実施の形態と実質的に同じ部分については同一の符号を付し、それらの説明を省略する。
【００５１】
図４の実施の形態における図１の実施の形態との違いは、入力軸１０がこれと平行に設けたカウンタ軸６２と第２歯車対５８を介して連結されており、このカウンタ軸６２上に第１遊星歯車組１４と第３クラッチ４６、第４クラッチ４８、第２ブレーキ５２およびワンウエイクラッチ５６が配置されるとともに、第１サンギヤ２０と第３サンギヤ３０とは第１チェーン６６で連結され、第１リングギヤ２２と第２キャリヤ３８とは第２チェーン６８で連結されていることである。
【００５２】
第２歯車対５８の一方の歯車は、第１歯車対４０の入力軸１０側の歯車が兼ねるようになっている。
なお、６６ａ、６６ｂ、６８ａ、６８ｂは第１チェーン６６および第２チェーン６８と噛み合うスプロケットを表す。
【００５３】
第１遊星歯車組１４がカウンタ軸６２上に配置された関係で、第３クラッチ４６は第１サンギヤとカウンタ軸６２とを連結し、第４クラッチ４８は第１キャリヤ２８とカウンタ軸６２とを連結するが、基本的に各締結要素の符号および作用は図１の実施の形態と同様であり、作動表は図２に示したものと共通である。
また、第２歯車対５８と第１チェーン６６および第２チェーン６８の変速比をそれぞれ１とすると、変速比も図１の実施の形態と同じになるので、詳細の説明を省略する。
【００５４】
図４に示した実施の形態も、前進８段後進２段の変速比が得られ、図１の実施の形態で説明したように、第３クラッチ４６を設けないと前進６段後進１段の変速比を得る歯車列にすることができる。
カウンタ軸６２側に配置されているのは、第１遊星歯車組１４と第１歯車対４０、第１チェーン６６、第２チェーン６８と第３クラッチ４６、第４クラッチ４８、第２ブレーキ５２およびワンウエイクラッチ５６であり、出力軸１２側に配置されているのは、第２遊星歯車組１６と第１歯車対４０、第１チェーン６６、第２チェーン６８と、第１クラッチ４２、第２クラッチ４４および第１ブレーキ５０である。
第２歯車対５８は第１歯車対４０と軸方向にオーバーラップしているので、歯車列の長さには影響しない。
【００５５】
また、変速機に一般的に用いる「はす歯歯車」は、動力伝達によって噛み合い部で軸方向のスラストが生じて、歯車を傾けるモーメントが作用するので、歯車を支持するベアリングが大きくなり勝ちであるが、チェーンの場合は軸方向のスラストが生じないため各スプロケット６６ａ、６６ｂ、６８ａ、６８ｂを支持するベアリングを簡素化できるので、軸方向の所要スペースを小さくてできるメリットがある。
【００５６】
これらの結果、図１に示した実施の形態と同様に、遊星歯車組と歯車対やチェーンおよび摩擦要素がカウンタ軸６２と駆動軸１６の両軸にバランスよく配置され、歯車列全体の軸方向長さを短くレイアウトした変速機が可能になるので、エンジン横置きの前輪駆動車に搭載する変速機への適用が容易になる。
【００５７】
次に、本発明の多段変速遊星歯車列における第３の実施の形態のスケルトンを図５に示す。
ここでは、図１および図４に示した実施の形態と異なる部分を中心に説明し、図１の実施の形態と実質的に同じ部分については同一の符号を付し、それらの説明を省略する。
【００５８】
図５の実施の形態は、図４に示した実施の形態と同様に、入力軸１０と平行に設けたカウンタ軸６２上に第１遊星歯車組１４が配置され、これらと平行に設けた駆動軸１６上に第２遊星歯車組１８が配置されているが、第２遊星歯車組１８がいわゆるダブルピニヨン型である。
【００５９】
すなわち、第２サンギヤ３０と、第２リングギヤ３２と、第２リングギヤ３２に噛み合った第２アウタピニヨン３４と、これら第２アウタピニヨン３４および第２サンギヤ３０に噛み合った第２インナピニヨン３６と、第２インナピニヨン３６および第２アウタピニヨン３４を回転自在に軸支する第２キャリヤ３８とで構成されている。
【００６０】
また、出力軸１２がカウンタ軸６２と同じ軸心上にあることも図４の実施の形態と異なる。
さらに、駆動軸１６およびカウンタ軸６２と第１、第２遊星歯車組１４、１８の回転メンバー間の連結関係が異なり、以下のようになっている。
駆動軸１６は、第１クラッチ４２により第２サンギヤ３０と、第２クラッチ４４により第２キャリヤ３８と、それぞれ選択的に連結可能である。
【００６１】
カウンタ軸６２は、第３クラッチ４６により第１サンギヤ２０と、第４クラッチ４８により第１キャリヤ２８と、それぞれ選択的に連結可能である。
第１サンギヤ２０は第１ブレーキ５０によりケース５４に固定可能である。
第２リングギヤ３２は第２ブレーキ５２によりケース５４に固定可能であるとともに、ワンウエイクラッチ５６により一方の回転方向において常にケース５４に固定されている。
【００６２】
第２キャリヤ３８と第１サンギヤ２０とは第１チェーン６６により連結されており、第１キャリヤ２８と第２リングギヤ３２とは第２チェーン６８により連結されている。
したがって、第１ブレーキ５０は第２キャリヤ３８を固定する機能も有し、第２ブレーキ５２およびワンウエイクラッチ５６は第１キャリヤ２８を固定する機能も有する。
第１リングギヤ２２は出力軸１２と連結されている。
【００６３】
上記のように各メンバー間の連結関係は異なるが、各締結要素の符号および作用は図１に示した実施の形態と同じであり、作動表も図２に示したものと共通である。
また、図３に示した共通速度線図においては、上記の連結関係を踏まえて各回転メンバーを置換すると同じように描くことができる。
【００６４】
すなわち、左端のＲ２の縦線に第２サンギヤ３０（Ｓ２）を、その右隣のＣ２、Ｒ１の縦線に第１リングギヤ２２（Ｒ１）を、さらに右隣のＣ１の縦線に第１キャリヤ２８（Ｃ１）および第２リングギヤ３２（Ｒ２）を、そして右端のＳ１、Ｓ２の縦線に第１サンギヤ２０（Ｓ１）および第２キャリヤ３８（Ｃ２）を割り当てることで各メンバーを表す縦線が同様になる。
【００６５】
したがって、第１遊星歯車組１４および第２遊星歯車組１８の歯数比α１、α２を適切に設定して共通速度線図の各縦線の間隔が図３に示したのと同じになるようにすると、変速比の値も図１に示した実施の形態と同じになる。
また、変速比の計算式自体は図１の実施の形態と異なるが、１／α２−１をＡと置いて上記の共通速度線図から導き出すことができるので説明を省略する。
【００６６】
図５に示した実施の形態も、前進８段後進２段の変速比が得られるとともに、
カウンタ軸６２側に配置されているのは、第１遊星歯車組１４と第１歯車対４０、第１、第２チェーン６６、６８と第３クラッチ４６、第４クラッチ４８および第１ブレーキ５０であり、駆動軸１６側に配置されているのは、第２遊星歯車組１６と第１歯車対４０、第１、第２チェーン６６、６８と、第１クラッチ４２、第２クラッチ４４、第２ブレーキ５２およびワンウエイクラッチ５６である。
【００６７】
したがって、カウンタ軸６２と駆動軸１６の両軸に略均等に分散して配置されることは図１に示した実施の形態と同様であり、歯車列全体の軸方向長さを短くレイアウトした変速機が可能になるので、エンジン横置きの前輪駆動車に搭載する変速機への適用が容易になる。
また、図１の実施の形態で説明したように、第３クラッチ４６を設けないで前進６段後進１段の変速比を得る構成にすることもできる。
【００６８】
次に、本発明の多段変速遊星歯車列における第４の実施の形態のスケルトンを図６に示す。
ここでは、図１および図５に示した実施の形態と異なる部分を中心に説明し、図１および図５に示した実施の形態と実質的に同じ部分については同一の符号を付し、それらの説明を省略する。
【００６９】
図６の実施の形態は、図５に示した実施の形態と基本的に同じである。
すなわち、違いは第１サンギヤ２０と第１チェーン６６を介して連結される相手が図５における第２キャリヤ３８から第２サンギヤ３０に替わったことである。
【００７０】
また、それと関連して第２キャリヤと駆動軸１６とを連結するクラッチが、図５の実施の形態における第２クラッチ４４から図６においては第１クラッチ４２に替わり、逆に第２サンギヤ３０と駆動軸１６とを連結するクラッチが図５における第１クラッチ４２から図６においては第２クラッチ４４に替わった。
つまり、図５に示した実施の形態とは第２サンギヤ３０と第２キャリヤ３８の役目が交替したことになる。
【００７１】
一般にダブルピニヨン型の遊星歯車組においてはサンギヤとキャリヤとが機能的に一方と他方の関係にあり、両者を逆にしても同じ機能を果たす。
したがって、図６に示した実施の形態は、基本的に図５に示した実施の形態と同じ作用をする。
【００７２】
各締結要素の符号および作用は図１および図５に示した実施の形態と同じであり、作動表も図２に示したものと共通である。
また、図３に示した共通速度線図においては、上記の連結関係を踏まえて各回転メンバーを置換すると同じように描くことができる。
【００７３】
すなわち、左端のＲ２の縦線に第２キャリヤ３８（Ｃ２）を、その右隣のＣ２、Ｒ１の縦線に第１リングギヤ２２（Ｒ１）を、さらに右隣のＣ１の縦線に第１キャリヤ２８（Ｃ１）および第２リングギヤ３２（Ｒ２）を、そして右端のＳ１、Ｓ２の縦線に第１サンギヤ２０（Ｓ１）および第２サンギヤ３０（Ｓ２）を割り当てることで各メンバーを表す縦線が同様になる。
【００７４】
したがって、第１遊星歯車組１４および第２遊星歯車組１８の歯数比α１、α２を適切に設定して共通速度線図の各縦線の間隔が図３に示したのと同じになるようにすると、変速比の値も図１に示した実施の形態と同じになる。
また、変速比の計算式自体は図１の実施の形態と異なるが、１／α２−１−α１をＡと置いて上記の共通速度線図から導き出すことができるので、その説明を省略する。
【００７５】
図６に示した実施の形態も、図５に示した実施の形態と同様に、前進８段後進２段の変速比が得られるとともに、カウンタ軸６２側および駆動軸１６側の両軸に略均等に分散して配置されることは図１に示した実施の形態と同様であり、歯車列全体の軸方向長さを短くレイアウトした変速機が可能になるので、エンジン横置きの前輪駆動車に搭載する変速機への適用が容易になる。
また、図１の実施の形態で説明したように、第３クラッチ４６を設けないで前進６段後進１段の変速比を得る構成にすることもできる。
【００７６】
以上、説明したように、本発明に係る多段変速遊星歯車列は、遊星歯車組や歯車対、チェーンおよび締結要素が、入力軸１０側もしくはカウンタ軸６２と駆動軸１６側の両軸に略均等に分散して配置できるのが特徴であり、従来例に比べて変速機の軸方向長さを短くレイアウトすることができる。
したがって、エンジン横置きの前輪駆動車に搭載する変速機への適用が容易になる。
【００７７】
車両の燃費や排気の浄化にとって、変速比が幅広くとれる多段変速機は必要不可欠であり、世界的に生産数量が多いエンジン横置きの前輪駆動車に適用可能な多段変速遊星歯車列は、省エネルギーおよび地球環境保全の観点からメリットが多い。
【００７８】
本発明に係る多段変速遊星歯車列によれば、上記したような効果が得られるほかに、さらに当業者の一般的な知識に基づいて、各締結要素のレイアウトや、歯車を支持するベアリングのレイアウトを工夫することや、第１ブレーキと並列に第２のワンウエイクラッチおよびこれをケースに固定するブレーキを追加するなどの改良や変更を加えた態様で実施することができる。
【図面の簡単な説明】
【図１】本発明の第１の実施の形態に係る多段変速遊星歯車列のスケルトンを示す図である。
【図２】図１に示した多段変速遊星歯車列の作動表を示す図である。
【図３】図１に示した多段変速遊星歯車列の共通速度線図および変速比の例を示す図である。
【図４】本発明の第２の実施の形態に係る多段変速遊星歯車列のスケルトンを示す図である。
【図５】本発明の第３の実施の形態に係る多段変速遊星歯車列のスケルトンを示す図である。
【図６】本発明の第４の実施の形態に係る多段変速遊星歯車列のスケルトンを示す図である。
【符号の説明】
１０：入力軸
１２：出力軸
１４：第１遊星歯車組
１６：駆動軸
１８：第２遊星歯車組
２０：第１サンギヤ
２２：第１リングギヤ
２４：第１ピニヨン
２８：第１キャリヤ
３０：第２サンギヤ
３２：第２リングギヤ
３４：第２ピニヨン、第２アウタピニヨン
３６：第２インナピニヨン
３８：第２キャリヤ
４０：第１歯車対
４２：第１クラッチ
４４：第２クラッチ
４６：第３クラッチ
４８：第４クラッチ
５０：第１ブレーキ
５２：第２ブレーキ
５４：ケース
５６：ワンウエイクラッチ
５８：第２歯車対
６０：第３歯車対
６２： カウンタ軸
６６： 第１チェーン
６８： 第２チェーン[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a multi-speed planetary gear train having a gear ratio of six forward speeds or more used for an automatic transmission for a vehicle.
[0002]
[Prior art]
As a conventional multi-speed planetary gear train having a gear ratio of 6 or more forward gears, two gear sets and two gear pairs are combined to obtain a gear ratio of 6 forward gears and 8 forward gears (for example, And Patent Document 1).
[0003]
[Patent Document 1]
JP 2002-323098 A
[0004]
[Problems to be solved by the invention]
The above conventional multi-speed planetary gear train has six friction elements (5 friction elements in the case of six forward speeds) on the same axis as the two planetary gear sets in order to obtain a forward gear ratio of eight speeds. If these are arranged, the required space in the axial direction becomes large. Therefore, it is difficult to apply to a transmission mounted on a so-called engine-mounted front-wheel drive vehicle in which the engine is disposed sideways. There was a problem.
[0005]
The present invention has been made in view of such conventional problems, and the axial length of the gear train is obtained while obtaining two or more planetary gear sets and a forward 6-speed or 8-speed gear ratio. It is an object of the present invention to improve the applicability of the engine horizontally mounted front wheel drive vehicle to a transmission.
[0006]
[Means for Solving the Problems]
In order to achieve the above object, a multi-stage planetary gear train according to the present invention described in claim 1 includes an input shaft, an output shaft disposed parallel to the input shaft, and a drive shaft disposed parallel to the input shaft. , Having a first planetary gear set and a second planetary gear set that are provided between the input shaft and the output shaft and convert the rotational speed of the input shaft to the rotational speed of the output shaft. A first sun gear, a first ring gear, a first pinion meshing with the first ring gear and the first sun gear, and a first carrier that pivotally supports the first pinion are disposed on the same axis as the shaft. A second planetary gear set is disposed on the drive shaft, and the second sun gear, the second ring gear, the second pinion meshed with the second ring gear and the second sun gear, and the second pinion A second carrier that pivotally supports the input A first gear pair connecting between the first and second drive gears, a second gear pair connecting between the first sun gear and the second sun gear, and a third gear connecting between the first ring gear and the second carrier. The first sun gear and the second sun gear can be fixed to the case, the second sun gear and the drive shaft can be connected, and either the first ring gear or the second carrier is the output shaft. The first carrier can be fixed to the case and can be connected to the input shaft, and the second ring gear can be connected to the drive shaft.
[0007]
In order to achieve the above object, a multi-speed planetary gear train according to a second aspect of the present invention includes an input shaft, an output shaft disposed in parallel to the input shaft, and a counter shaft disposed in parallel to the input shaft. A first planetary gear set and a second planetary gear set which are provided between the input shaft and the output shaft, and convert the rotational speed of the input shaft into the rotational speed of the output shaft. The first planetary gear set is disposed on the same axis as the counter shaft, and the first sun gear, the first ring gear, and the first ring gear and the first sun gear are engaged with each other. A pinion and a first carrier that pivotally supports the first pinion, the second planetary gear set being disposed coaxially with the drive shaft, a second sun gear, a second ring gear, and the first gear 2nd gear meshing with 2 ring gear and 2nd sun gear And a second carrier for supporting the second pinion, a first gear pair for connecting the input shaft and the drive shaft, and a second for connecting the input shaft and the counter shaft. A gear pair, a first chain that connects the first sun gear and the second sun gear, and a second chain that connects the first ring gear and the second carrier are provided. The sun gear can be fixed to the case, the second sun gear and the counter shaft can be connected, either the first ring gear or the second carrier can be connected to the output shaft, and the first carrier can be fixed to the case In addition, the second ring gear can be connected to the drive shaft and can be connected to the counter shaft.
[0008]
In order to achieve the above object, the multi-speed planetary gear train of the present invention described in claim 3 is characterized in that the first sun gear and the input shaft or the counter shaft can be connected.
[0009]
In order to achieve the above object, a multi-speed planetary gear train according to a fourth aspect of the present invention includes an input shaft, an output shaft disposed parallel to the input shaft, and a counter shaft disposed parallel to the input shaft. A first planetary gear set and a second planetary gear set which are provided between the input shaft and the output shaft, and convert the rotational speed of the input shaft into the rotational speed of the output shaft. A planetary gear set, the first planetary gear set being arranged coaxially with the counter shaft, and a first sun gear, a first ring gear, and a first pinion meshing with the first ring gear and the first sun gear; A first carrier that pivotally supports the first pinion, and a second planetary gear set is disposed coaxially with the drive shaft, and a second sun gear, a second ring gear, and the second ring gear. A second outer pinion meshing with the second outer pinion, A first gear that includes a second inner pinion that meshes with the outer pinion and the second sun gear, and a second carrier that pivotally supports the second inner pinion and the second outer pinion, and that connects the input shaft and the drive shaft. A pair of second gears connecting between the input shaft and the counter shaft; a first chain connecting between the first sun gear and the second carrier or the second sun gear; and a first carrier and a second ring gear. A first chain and the second carrier or the second sun gear can be fixed to the case, the second carrier and the drive shaft can be connected, and the first carrier And the second ring gear can be fixed to the case, the first carrier and the counter shaft can be connected, the first ring gear can be connected to the output shaft, and the second sun gear can be connected to the drive shaft. And characterized in that.
[0010]
[Action and effect]
In the multi-speed planetary gear train of the present invention described in claim 1, the first planetary gear set is arranged coaxially with the input shaft, the second planetary gear set is arranged on the drive shaft, and the input shaft A first gear pair for connecting the drive shaft, a second gear pair for connecting the first sun gear and the second sun gear, and a third gear pair for connecting the first ring gear and the second carrier. The first sun gear and the second sun gear can be fixed to the case, the second sun gear and the drive shaft can be connected, and either the first ring gear or the second carrier can be connected to the output shaft. Since the first carrier can be fixed to the case and can be connected to the input shaft, and the second ring gear can be connected to the drive shaft, when the drive shaft and the second ring gear are connected, other members can be connected. In combination with the relationship at least the first forward speed Performs driving up to the fourth speed, to drive the high-speed stage on at least a fifth forward gear in combination with the connection of other members when connecting the input shaft and the first carrier.
As a result, the planetary gear set, the gear pair, and the friction element can be arranged substantially evenly distributed on the input shaft side and the drive shaft side while obtaining a forward gear ratio of 6 or 8 speeds. A shorter layout becomes possible, and the application to the transmission of the engine-side-mounted front-wheel drive vehicle becomes easy.
[0011]
The multi-speed planetary gear train according to the present invention described in claim 2 has a counter shaft and a drive shaft arranged in parallel to the input shaft, and the first planetary gear set is on the same axis as the counter shaft. The second planetary gear set is arranged coaxially with the drive shaft, the first gear pair connecting the input shaft and the drive shaft, and the second gear connecting the input shaft and the counter shaft And a first chain that connects the first sun gear and the second sun gear; and a second chain that connects the first ring gear and the second carrier, the first sun gear and the second sun gear. Can be fixed to the case, the second sun gear and the counter shaft can be connected, one of the first ring gear and the second carrier can be connected to the output shaft, and the first carrier can be fixed to the case. The second ring can be connected to the counter shaft Since the drive shaft can be connected to the drive shaft, when the drive shaft and the second ring gear are connected, the drive from the first forward speed to the fourth speed or the fifth speed is performed in combination with the connection relationship of the other members. When the first carrier is connected to the first carrier, the high speed stage is driven from the fourth forward speed or the fifth speed forward in combination with the connection relationship of the other members.
As a result, the planetary gear set, the gear pair, the chain, and the friction element can be arranged almost evenly distributed on the counter shaft side and the drive shaft side, so that a layout with a reduced axial length is possible. Application to a transmission of a front-wheel drive type vehicle becomes easy.
[0012]
In the multi-speed planetary gear train of the present invention described in claim 3, since the first sun gear can be connected to the input shaft or the counter shaft, the first sun gear is connected to the input shaft, When the second ring gear is connected, the fourth forward speed is driven, and when the input shaft and the first carrier are connected, the sixth forward speed is driven.
As a result, the planetary gear set, the gear pair, and the friction element can be arranged in a balanced manner on the input shaft side and the drive shaft side while obtaining a gear ratio of 8 forward speeds and 2 reverse speeds. It becomes possible, and the application to the transmission of the engine horizontal type front wheel drive vehicle becomes easy.
[0013]
In the multi-speed planetary gear train of the present invention described in claim 4, the counter shaft and the drive shaft arranged in parallel with the input shaft, the first planetary gear set is arranged coaxially with the counter shaft, Two planetary gear sets are arranged coaxially with the drive shaft, a first gear pair that connects the input shaft and the drive shaft, a second gear pair that connects the input shaft and the counter shaft, A first chain connecting between the sun gear and the second carrier or the second sun gear; and a second chain connecting between the first carrier and the second ring gear, the first sun gear and the second carrier or The second sun gear can be fixed to the case, the second carrier and the drive shaft can be connected, the first carrier and the second ring gear can be fixed to the case, and the first carrier and the counter shaft can be connected. Is possible and the first ring gear is Since the second sun gear is configured to be connectable to the drive shaft and connected to the force shaft, when the drive shaft and the second sun gear or the second carrier are connected, it is combined with the connection relationship of the other members to at least the first forward speed from the first forward speed. Driving up to the fourth speed is performed, and when the counter shaft and the first carrier are connected, the high speed stage is driven at least from the fifth forward speed in combination with the connection relationship of the other members.
As a result, the planetary gear set, the gear pair, the chain, and the friction element can be arranged substantially evenly distributed on the counter shaft side and the drive shaft side while obtaining a forward gear ratio of 6 or 8 steps. A shorter layout is possible, and the application to the transmission of a horizontally-mounted front-wheel drive vehicle becomes easy.
[0014]
DETAILED DESCRIPTION OF THE INVENTION
Embodiments of the multi-speed planetary gear train of the present invention will be described below with reference to the drawings.
FIG. 1 is a skeleton diagram showing a multi-speed planetary gear train according to the first embodiment of the present invention.
In the figure, of the input shaft 10 and the output shaft 12 arranged in parallel to each other, the input shaft 10 side shows the lower half of the axis, and the output shaft 12 side shows the upper half of the axis.
[0015]
In this multi-stage planetary gear train, the input shaft 10 and the output shaft 12 are arranged in parallel to each other as described above so as to be applied to a transmission of a front wheel drive vehicle in which an engine (not shown) is mounted sideways. A first planetary gear set 14 is arranged on the same axis as 10, and a second planetary gear set 18 is arranged on a drive shaft 16 having the same axis as the output shaft 12.
[0016]
The first planetary gear set 14 and the second planetary gear set 18 are generally called single pinion types, and each has the same configuration. The first planetary gear set 14 rotatably supports the first sun gear 20, the first ring gear 22, the first pinion 24 meshed with the first ring gear 22 and the first sun gear 20, and the first pinion 24. The first carrier 28 is configured.
Similarly, the second planetary gear set 18 includes a second sun gear 30, a second ring gear 32, a second pinion 34, and a second carrier 38.
[0017]
The input shaft 10, the output shaft 12, the drive shaft 16, and the rotating members of the first planetary gear set 14 and the second planetary gear set 18 are connected or connectable as follows.
The drive shaft 16 is connected to the input shaft 10 by the first gear pair 40 and is driven at a reduced speed.
The drive shaft 16 can be selectively connected to the second ring gear 32 by the first clutch 42 and the second sun gear 30 by the second clutch 44.
[0018]
The input shaft 10 can be selectively connected to the first sun gear 20 by the third clutch 46 and the first carrier 28 by the fourth clutch 48.
The second sun gear 30 can be fixed to the case (stationary part) 54 by the first brake 50. The first carrier 28 can be fixed to the case 54 by the second brake 52 and is always fixed to the case 54 in one rotational direction by the one-way clutch 56.
[0019]
The first sun gear 20 and the second sun gear 30 are connected by a second gear pair 58.
The first ring gear 22 and the second carrier 38 are connected by a third gear pair 60, and the second carrier 38 is connected to the output shaft 12.
Reference numeral 12a denotes an output gear.
[0020]
Next, the operation of the multi-speed planetary gear train shown in FIG. 1 will be described with reference to the operation table shown in FIG. 2 and the common speed diagram shown in FIG.
In the following description, clutches and brakes are referred to as friction elements, and those having a function of connecting rotating members including a one-way clutch are collectively referred to as fastening elements.
[0021]
In the operation table of FIG. 2, fastening elements such as a clutch, a brake, and a one-way clutch are assigned to the horizontal column, C-1 represents the first clutch 42, B-1 represents the first brake 50, OC represents the one-way clutch 56 and so on. The relationship between these symbols and the symbols of the respective fastening elements is shown in FIG.
[0022]
The vertical column is divided into “D range”, “R range” and “L range” of the operation lever (not shown), and the D range is assigned with the forward first speed (1st) to the eighth speed (8th), and the R range. Are assigned to the first reverse speed (R-1) and the second speed (R-2).
In the L range, it is possible to drive the input shaft 10 side from the output shaft 12 side as in engine braking described later.
In the operation table of FIG. 2, “o” indicates the fastening of each fastening element, and the blank indicates the release of each fastening element.
[0023]
As shown in FIG. 3, in the common speed diagram of the multi-speed planetary gear train of FIG. 1, the vertical direction represents the rotational speed of each rotating member when the rotational speed of the input shaft 10 is 1, and the horizontal direction represents the first rotational speed. The rotating members are allocated at intervals corresponding to the gear ratio of the first and second planetary gear sets 14 and 18, and a vertical line is drawn for each rotating member.
[0024]
The symbol written above the common speed diagram is S for the sun gear, R for the ring gear, C for the carrier, and the subsequent numbers 1 and 2 indicate the first and second planetary gear sets to which the gear belongs, for example, S1 , R1 and C1 represent the first sun gear 20, the first ring gear 22, and the first carrier 28 of the first planetary gear set 14, respectively.
[0025]
Here, the gear ratio of each planetary gear set is a ratio Zs / Zr of the number of teeth of the sun gear (Zs) to the number of teeth of the ring gear (Zr), and the gear ratio of the first planetary gear set 14 is α1, The gear ratio of the two planetary gear set 18 is α2.
The rotation speed of the second ring gear 32 (R2) in the common speed diagram is represented by i1 as the reduction ratio of the first gear pair 40.
[0026]
In order to simplify the calculation and explanation of the gear ratio, both the gear ratio i2 of the second gear pair 58 and the gear ratio i3 of the third gear pair 60 are set to 1.
Accordingly, in the common speed diagram, the first ring gear 22 (R1) and the second carrier 38 (C2), the first sun gear 20 (S1), and the second sun gear 30 (S2) have a common vertical line.
[0027]
In the calculation of the gear ratio including the common speed diagram, a case where α1 is 0.50, α2 is 0.45, and i1 is 2.00 will be described.
Therefore, the drive shaft 16 rotates at half the rotational speed of the input shaft 10.
Furthermore, in order to simplify the display and calculation formula, α2 (1 + α1) is A.
[0028]
In the common velocity diagram, the height of the intersection of each vertical line and the thick line represents the number of rotations of each rotation member.
For easy understanding, the intersections of the second carrier 38 (C2) connected to the output shaft 12 on the vertical line are indicated by circles.
FIG. 3B shows the gear ratios and the inter-step ratios when α1, α2, and i3 are set to the above values.
[0029]
First, the forward first speed (1st) is driven by connecting the second ring gear 32 to the drive shaft 16 by engaging the first clutch 42 (C-1) as seen in the operation table shown in FIG. Done.
At this time, the first carrier 28 is fixed to the case 54 by a one-way clutch 56 (OC).
[0030]
That is, the one-way clutch 56 is at the first speed, and the first carrier 28 is fixed to the case 54 in the direction of accelerating the vehicle. At the first speed in the D range, the output shaft is at the time of so-called engine braking. No drive from 12 to the input shaft 10 is possible.
[0031]
The speed ratio of the first speed (the rotational speed of the input shaft 10 / the rotational speed of the output shaft 12) is i1 (1 + α2) + i1 · α2 / α1, and the gear ratio set to the above value is 4.700. Become.
This calculation formula is a case where i2 and i3 are set to 1 as described above, and when converted to a formula including i2 and i3 for reference, i1 (1 + α2) + i1 · i3 · α2 / (α1 · i2) is obtained. .
Here, i2 and i3 are Zd / Zi where the number of teeth on the drive shaft 16 side is Zd and the number of teeth on the input shaft 10 side is Zi.
[0032]
The first speed will be described with reference to the common speed diagram of FIG. 3. Assuming that the rotation speed of the input shaft 10 is 1, the second ring gear 32 (R2) is 1 / i1 and the first carrier 28 (C1) is the case. Since the rotational speed is 0, the height of the intersection of the 1st diagonal line connecting the two and the vertical line of the second carrier 38 (C2) is the rotational speed of the output shaft 12.
[0033]
Next, the speed change to the second speed (2nd) is performed by engaging the first brake 50 (B-1) in addition to the engagement of the first clutch 42 at the first speed described above. This is done by fixing to the case 54.
At this time, the fixing of the first carrier 28 to the case 54 is automatically released by the action of the one-way clutch 56.
[0034]
By fixing the second sun gear 30 to the case 54, the output shaft 12 is driven to decelerate from the drive shaft 16 by the second planetary gear set 18, and the rotation speed of the second sun gear 30 (S2) is shown in the common speed diagram. A 2nd diagonal line with 0 is shown, and the speed ratio of the second speed is i1 (1 + α2).
In the above-mentioned tooth number ratio, it is 2.900.
[0035]
As described above, the shift from the first forward speed to the second speed has only the addition of the first brake 50 since the one-way clutch 56 has an action.
Therefore, a so-called shift shock at the time of shifting can be suppressed only by performing control so that the first brake 50 is gently engaged, so that smooth shift control can be easily performed.
[0036]
Next, the shift to the third speed (3rd) is performed by releasing the engagement of the first brake 50 at the second speed and engaging the second clutch 44.
As a result, the second planetary gear set 18 is united, the 3rd horizontal line is shown in the common speed diagram, and the first planetary gear pair 40 is decelerated and driven, and the gear ratio is i1. Become.
In the above-mentioned tooth number ratio, it is 2.000.
[0037]
Subsequently, the shift to the fourth speed (4th) is performed by releasing the engagement of the second clutch 44 at the third speed and engaging the third clutch 46.
In the common speed diagram, a 4th diagonal line with the rotation speed of the second sun gear 30 (S2) being 1 is shown, and the speed ratio of the fourth speed is i1 (1 + A + α1) / (1 + α1 + A · i1). .
In the above-mentioned tooth number ratio, it is 1.526.
[0038]
Next, the shift to the fifth speed (5th) is performed by releasing the engagement of the third clutch 46 at the fourth speed and engaging the fourth clutch 48.
The engagement of the fourth clutch 48 is maintained at the subsequent high speed.
As a result, the first carrier 28 is connected to the input shaft 10 and becomes a 5th oblique line with the rotation speed of the first carrier 28 (C1) being 1 in the common speed diagram, and the transmission ratio is i1 (A + α1) / (A I1 + α1)
In the above-mentioned tooth number ratio, it becomes 1.270.
[0039]
Next, the shift to the sixth speed (6th) is performed by releasing the engagement of the first clutch 42 up to the fifth speed and engaging the third clutch 46 again.
At this stage, as shown by the 6th horizontal line in the common speed diagram, the first planetary gear set 14 is integrated with the input shaft 10, the output shaft 12 is driven only by the third gear pair 60, and the transmission ratio is i3. Therefore, the above-mentioned tooth number ratio is 1.000.
[0040]
Next, in the shift to the seventh speed (7th), the engagement of the third clutch 46 at the sixth speed is released, and the second clutch 44 is engaged again to connect the second sun gear 30 to the drive shaft 16. Is done.
As a result, as shown in the common speed diagram, the 1st carrier 28 (C1) has a rotational speed of 1 and the second sun gear 30 (S2) has a rotational speed of 1 / i1, resulting in a 7th diagonal line, and the gear ratio. Becomes i1 / {i1 + α1 (i1-1)}, and the speed increase (overdrive) is 0.800 in the above-described ratio of the number of teeth.
[0041]
Next, in shifting to the eighth speed (8th), the second clutch 44 in the seventh speed is released and the first brake 50 is engaged again to fix the first sun gear 20 to the case 54. Done in
Actually, the first brake 50 is fixed to the second sun gear 30, but the second sun gear 30 and the first sun gear 20 are connected by the second gear pair 58, so the first sun gear 20 is also fixed. It will be.
As a result, as shown in the common speed diagram, the first carrier 28 (C1) has a rotational speed of 1 and the first sun gear 20 (S1) has a rotational speed of 0, resulting in an 8th diagonal line, and the transmission ratio is i1. / (1 + α1), and the speed increase is 0.667 in the above-described ratio of teeth.
[0042]
Next, the reverse first speed (R-1) of the R range is driven by engaging the second clutch 44 and the second brake 52.
As a result, the first sun gear 20 is driven from the drive shaft 16 and the first carrier 28 is fixed to the case 54. Therefore, the first sun gear 20 is reversed and drives the output shaft 12 via the third gear pair 60.
In the common speed diagram, the rotation speed of the first sun gear 20 (S1) is 1 / i1, and the rotation speed of the first carrier 28 (C1) is 0, as indicated by the diagonal line of R-1, and the gear ratio is -I1 / α1, and in the above-mentioned ratio of the number of teeth, becomes -4,000.
[0043]
Further, the shift to the second reverse speed (R-2) is performed by releasing the engagement of the second clutch 44 at the first reverse speed and engaging the third clutch 46.
As a result, the first sun gear 20 is directly connected to the input shaft 10 and rotates at the rotation speed 1, and in the common speed diagram, the first sun gear 20 (S1) has a rotation speed of 1, and is indicated by the diagonal line R-2. Thus, the gear ratio becomes -1 / α1, and in the above-mentioned gear ratio, it becomes -2,000.
[0044]
As described above, since the one-way clutch 56 is automatically engaged only in the direction of accelerating the vehicle at the first speed in the D range, when driving from the output shaft 12 side during engine braking, FIG. As shown in 1st in the L range of the operation table, the first clutch 42 is fastened in the same manner as the first forward speed of the D range, and the second brake 52 provided with the one-way clutch 56 is fastened.
Thereby, the gear ratio of the forward first speed can be obtained regardless of the direction in which the torque acts.
[0045]
The above transmission ratio is summarized in FIG.
The ratio between the adjacent gear ratios is the interstage ratio.
As can be seen, the interstage ratio generally decreases with increasing speed toward the high speed stage, and this tends to be preferable as the transmission ratio of a vehicle transmission driven by an internal combustion engine.
[0046]
The above is the operation and the gear ratio in the multi-stage planetary gear train of eight forward speeds and two reverse speeds shown in FIG.
As described above, the first planetary gear set 14, the first to third gear pairs 40, 58, 60, the third clutch 46, the fourth clutch 48, and the second brake are arranged on the input shaft 10 side. 52 and the one-way clutch 56, which are disposed on the output shaft 12 side, are the second planetary gear set 16, the three gear pairs 40, 58, 60, the first clutch 42, the second clutch 44, and This is the first brake 50.
[0047]
In other words, the planetary gear set, the gear pair, and the friction element are arranged on the two shafts in a substantially uniform manner, and the gear pair is compared with a general transmission in which the input shaft and the output shaft are arranged in parallel. However, it has a great effect in laying out the entire axial length of the gear train as short as possible.
As a result, a transmission having a shorter axial length than that of the conventional example can be achieved, and therefore, it can be easily applied to a transmission mounted on a front-wheel drive vehicle mounted horizontally on the engine.
The first brake 50 that fixes the second sun gear 30 to the case 54 has the same effect even if it is arranged on the input shaft 10 side and fixed to the first sun gear 20. Arrangement can be freely selected.
[0048]
Further, in FIG. 1, the second carrier 38 is connected to the output shaft 12, but the output shaft 12 can be moved to the same axial center side as the input shaft 10 to be connected to the first ring gear 22.
If the third clutch 46 is omitted, the fourth forward speed, the sixth speed, and the second reverse speed in the operation table shown in FIG. Obtainable.
[0049]
In this case, for example, if α1 is 0.48, α2 is 0.58, and i1 is 1.5, the fourth speed, the sixth speed, and the second reverse speed in the above formula are deleted and the forward first speed is increased. When rearranged from the 6th gear to the 6th gear, the gear ratios are as follows.

Needless to say, although it has six forward speeds, a gear ratio suitable for a vehicle transmission can be obtained, and the axial length can be made shorter than in the conventional example.
[0050]
Next, FIG. 4 shows a skeleton of the second embodiment in the multi-speed planetary gear train of the present invention.
Here, parts different from the embodiment shown in FIG. 1 will be mainly described, and parts substantially the same as those of the embodiment of FIG. 1 are denoted by the same reference numerals, and description thereof will be omitted.
[0051]
The difference between the embodiment of FIG. 4 and the embodiment of FIG. 1 is that the input shaft 10 is connected to a counter shaft 62 provided in parallel with the input shaft 10 via a second gear pair 58. The first planetary gear set 14, the third clutch 46, the fourth clutch 48, the second brake 52, and the one-way clutch 56 are disposed, and the first sun gear 20 and the third sun gear 30 are connected by a first chain 66. The first ring gear 22 and the second carrier 38 are connected by a second chain 68.
[0052]
One gear of the second gear pair 58 serves as the gear on the input shaft 10 side of the first gear pair 40.
Reference numerals 66 a, 66 b, 68 a, and 68 b denote sprockets that mesh with the first chain 66 and the second chain 68.
[0053]
With the first planetary gear set 14 disposed on the counter shaft 62, the third clutch 46 connects the first sun gear and the counter shaft 62, and the fourth clutch 48 connects the first carrier 28 and the counter shaft 62. Although connected, the sign and operation of each fastening element are basically the same as those in the embodiment of FIG. 1, and the operation table is the same as that shown in FIG.
Further, assuming that the gear ratio of the second gear pair 58 and the first chain 66 and the second chain 68 is 1, the gear ratio is the same as that of the embodiment of FIG.
[0054]
In the embodiment shown in FIG. 4 as well, a transmission gear ratio of 8 forward speeds and 2 reverse speeds can be obtained. As described in the embodiment of FIG. A gear train for obtaining a gear ratio can be obtained.
Arranged on the counter shaft 62 side are the first planetary gear set 14 and the first gear pair 40, the first chain 66, the second chain 68 and the third clutch 46, the fourth clutch 48, the second brake 52, and The one-way clutch 56, which is disposed on the output shaft 12 side, is the second planetary gear set 16, the first gear pair 40, the first chain 66, the second chain 68, the first clutch 42, the second clutch. 44 and the first brake 50.
Since the second gear pair 58 overlaps the first gear pair 40 in the axial direction, the length of the gear train is not affected.
[0055]
In addition, the “helical gear” generally used in transmissions has an axial thrust at the meshing portion due to power transmission, and a moment that tilts the gear acts. Therefore, the bearing that supports the gear tends to be large. However, in the case of a chain, since axial thrust does not occur, the bearings supporting the sprockets 66a, 66b, 68a, 68b can be simplified, and there is an advantage that the required space in the axial direction can be reduced.
[0056]
As a result, similar to the embodiment shown in FIG. 1, the planetary gear set, the gear pair, the chain, and the friction element are arranged in a balanced manner on both the counter shaft 62 and the drive shaft 16, and the axial direction of the entire gear train Since a transmission with a short layout is possible, it can be easily applied to a transmission mounted on a front-wheel drive vehicle with an engine installed horizontally.
[0057]
Next, FIG. 5 shows a skeleton of the third embodiment in the multi-speed planetary gear train of the present invention.
Here, the description will focus on parts that are different from the embodiment shown in FIGS. 1 and 4, and parts that are substantially the same as those in the embodiment of FIG. .
[0058]
In the embodiment shown in FIG. 5, the first planetary gear set 14 is arranged on the counter shaft 62 provided in parallel with the input shaft 10 and the drive provided in parallel with these as in the embodiment shown in FIG. The second planetary gear set 18 is disposed on the shaft 16, and the second planetary gear set 18 is a so-called double pinion type.
[0059]
That is, the second sun gear 30, the second ring gear 32, the second outer pinion 34 meshed with the second ring gear 32, the second inner pinion 36 meshed with the second outer pinion 34 and the second sun gear 30, and the second inner The pinion 36 and the second outer pinion 34 are constituted by a second carrier 38 that rotatably supports the pinion 36 and the second outer pinion 34.
[0060]
Also, the output shaft 12 is on the same axis as the counter shaft 62, which is different from the embodiment of FIG.
Further, the connection relationship between the drive shaft 16 and the counter shaft 62 and the rotating members of the first and second planetary gear sets 14 and 18 is different, and is as follows.
The drive shaft 16 can be selectively connected to the second sun gear 30 by the first clutch 42 and the second carrier 38 by the second clutch 44.
[0061]
The counter shaft 62 can be selectively connected to the first sun gear 20 by the third clutch 46 and the first carrier 28 by the fourth clutch 48.
The first sun gear 20 can be fixed to the case 54 by the first brake 50.
The second ring gear 32 can be fixed to the case 54 by the second brake 52 and is always fixed to the case 54 in one rotational direction by the one-way clutch 56.
[0062]
The second carrier 38 and the first sun gear 20 are connected by a first chain 66, and the first carrier 28 and the second ring gear 32 are connected by a second chain 68.
Therefore, the first brake 50 also has a function of fixing the second carrier 38, and the second brake 52 and the one-way clutch 56 also have a function of fixing the first carrier 28.
The first ring gear 22 is connected to the output shaft 12.
[0063]
As described above, the connection relationship between the members is different, but the signs and actions of the fastening elements are the same as those in the embodiment shown in FIG. 1, and the operation table is also the same as that shown in FIG.
Further, in the common velocity diagram shown in FIG. 3, it can be drawn in the same manner when each rotating member is replaced in consideration of the above connection relation.
[0064]
That is, the second sun gear 30 (S2) is located on the leftmost vertical line of R2, the first ring gear 22 (R1) is located on the right next to C2, and the vertical line of R1, and the first carrier is located on the right next to the vertical line of C1. 28 (C1) and the second ring gear 32 (R2), and the first sun gear 20 (S1) and the second carrier 38 (C2) are assigned to the vertical lines S1 and S2 at the right end so that the vertical lines representing the members are It will be the same.
[0065]
Therefore, the tooth number ratios α1 and α2 of the first planetary gear set 14 and the second planetary gear set 18 are appropriately set so that the intervals between the vertical lines in the common speed diagram are the same as those shown in FIG. Then, the value of the gear ratio is the same as that of the embodiment shown in FIG.
Further, although the calculation formula for the transmission ratio itself is different from that of the embodiment of FIG. 1, it can be derived from the above-mentioned common speed diagram by setting 1 / α2-1 as A, and thus the description thereof is omitted.
[0066]
In the embodiment shown in FIG. 5 as well, a gear ratio of 8 forward speeds and 2 reverse speeds can be obtained.
Arranged on the counter shaft 62 side are the first planetary gear set 14 and the first gear pair 40, the first and second chains 66 and 68, the third clutch 46, the fourth clutch 48 and the first brake 50. The second planetary gear set 16 and the first gear pair 40, the first and second chains 66 and 68, the first clutch 42, the second clutch 44, and the second are arranged on the drive shaft 16 side. A brake 52 and a one-way clutch 56 are provided.
[0067]
Therefore, it is the same as the embodiment shown in FIG. 1 that the counter shaft 62 and the drive shaft 16 are arranged so as to be distributed almost evenly. This makes it possible to easily apply the transmission to a transmission mounted on a front-wheel drive vehicle with a horizontal engine.
Further, as described in the embodiment of FIG. 1, it is also possible to obtain a transmission ratio of 6 forward speeds and 1 reverse speed without providing the third clutch 46.
[0068]
Next, FIG. 6 shows a skeleton of the fourth embodiment in the multi-speed planetary gear train of the present invention.
Here, the description will focus on the parts different from the embodiment shown in FIG. 1 and FIG. 5, and the substantially same parts as those in the embodiment shown in FIG. 1 and FIG. The description of is omitted.
[0069]
The embodiment shown in FIG. 6 is basically the same as the embodiment shown in FIG.
That is, the difference is that the counterpart connected to the first sun gear 20 via the first chain 66 is changed from the second carrier 38 to the second sun gear 30 in FIG.
[0070]
In addition, the clutch that connects the second carrier and the drive shaft 16 in relation to the second clutch 44 in the embodiment of FIG. 5 is replaced with the first clutch 42 in FIG. The clutch that connects the drive shaft 16 is changed from the first clutch 42 in FIG. 5 to the second clutch 44 in FIG. 6.
That is, the role of the second sun gear 30 and the second carrier 38 is changed from the embodiment shown in FIG.
[0071]
In general, in a double pinion type planetary gear set, a sun gear and a carrier are functionally in a relationship between one and the other, and perform the same function even if they are reversed.
Therefore, the embodiment shown in FIG. 6 basically operates in the same manner as the embodiment shown in FIG.
[0072]
The sign and action of each fastening element are the same as those of the embodiment shown in FIGS. 1 and 5, and the operation table is also the same as that shown in FIG.
Further, in the common velocity diagram shown in FIG. 3, it can be drawn in the same manner when each rotating member is replaced in consideration of the above connection relation.
[0073]
That is, the second carrier 38 (C2) is placed on the leftmost vertical line of R2, the first ring gear 22 (R1) is placed on the vertical line of C2 and R1 on the right, and the first carrier is placed on the vertical line of C1 on the right. 28 (C1) and the second ring gear 32 (R2), and the first sun gear 20 (S1) and the second sun gear 30 (S2) are assigned to the vertical lines of S1 and S2 at the right end so that the vertical lines representing the members are It will be the same.
[0074]
Therefore, the tooth number ratios α1 and α2 of the first planetary gear set 14 and the second planetary gear set 18 are appropriately set so that the intervals between the vertical lines in the common speed diagram are the same as those shown in FIG. Then, the value of the gear ratio is the same as that of the embodiment shown in FIG.
Further, although the transmission ratio calculation formula itself is different from that of the embodiment of FIG. 1, it can be derived from the above-mentioned common speed diagram by setting 1 / α2-1-α1 as A, and the description thereof will be omitted.
[0075]
In the embodiment shown in FIG. 6, as in the embodiment shown in FIG. 5, a transmission gear ratio of 8 forward speeds and 2 reverse speeds is obtained, and both the counter shaft 62 side and the drive shaft 16 side are substantially omitted. 1 is the same as the embodiment shown in FIG. 1, and a transmission having a layout in which the axial length of the entire gear train is shortened is possible. Application to a transmission mounted on the
Further, as described in the embodiment of FIG. 1, it is also possible to obtain a transmission ratio of 6 forward speeds and 1 reverse speed without providing the third clutch 46.
[0076]
As described above, in the multi-speed planetary gear train according to the present invention, the planetary gear set, the gear pair, the chain, and the fastening element are substantially equal to both the input shaft 10 side or the counter shaft 62 and the drive shaft 16 side. As a feature, the transmission can be arranged in a distributed manner, and the axial length of the transmission can be laid out shorter than in the conventional example.
Therefore, it becomes easy to apply to a transmission mounted on a front-wheel drive vehicle with the engine placed horizontally.
[0077]
A multi-stage transmission with a wide gear ratio is indispensable for vehicle fuel efficiency and exhaust purification, and a multi-speed planetary gear train that can be applied to front-wheel-drive vehicles with horizontal engine production, which has a large production volume around the world. There are many advantages from the viewpoint of global environmental conservation.
[0078]
According to the multi-speed planetary gear train according to the present invention, in addition to the above-described effects, the layout of the fastening elements and the layout of the bearings supporting the gears are further based on the general knowledge of those skilled in the art. It is possible to carry out the invention in a manner in which improvements or changes are made, such as adding a second one-way clutch and a brake for fixing the second one-way clutch to the case in parallel with the first brake.
[Brief description of the drawings]
FIG. 1 is a diagram showing a skeleton of a multi-speed planetary gear train according to a first embodiment of the present invention.
FIG. 2 is a diagram showing an operation table of the multi-speed planetary gear train shown in FIG. 1;
FIG. 3 is a diagram showing a common speed diagram and an example of a gear ratio of the multi-speed planetary gear train shown in FIG. 1;
FIG. 4 is a diagram showing a skeleton of a multi-speed planetary gear train according to a second embodiment of the present invention.
FIG. 5 is a diagram showing a skeleton of a multi-speed planetary gear train according to a third embodiment of the present invention.
FIG. 6 is a diagram showing a skeleton of a multi-speed planetary gear train according to a fourth embodiment of the present invention.
[Explanation of symbols]
10: Input shaft
12: Output shaft
14: First planetary gear set
16: Drive shaft
18: Second planetary gear set
20: First sun gear
22: First ring gear
24: First pinion
28: First carrier
30: Second sun gear
32: Second ring gear
34: 2nd pinion, 2nd outer pinion
36: Second inner pinion
38: Second carrier
40: First gear pair
42: First clutch
44: Second clutch
46: Third clutch
48: Fourth clutch
50: First brake
52: Second brake
54: Case
56: One-way clutch
58: Second gear pair
60: Third gear pair
62: Counter axis
66: First chain
68: Second chain

Claims (4)

An input shaft;
An output shaft disposed parallel to the input shaft;
A drive shaft disposed parallel to the input shaft;
A first planetary gear set and a second planetary gear set which are provided between the input shaft and the output shaft and convert the rotational speed of the input shaft into the rotational speed of the output shaft;
The first planetary gear set is disposed coaxially with the input shaft, and includes a first sun gear, a first ring gear, a first pinion meshed with the first ring gear and the first sun gear, and the first A first carrier that pivotally supports the pinion,
The second planetary gear set is disposed on the drive shaft, and a second sun gear, a second ring gear, a second pinion engaged with the second ring gear and the second sun gear, and the second pinion A second carrier that pivotally supports,
A first gear pair connecting between the input shaft and the drive shaft;
A second gear pair connecting between the first sun gear and the second sun gear;
A third gear pair that connects between the first ring gear and the second carrier;
The first sun gear and the second sun gear can be fixed to a case;
The second sun gear and the drive shaft are connectable;
Either the first ring gear or the second carrier is connected to the output shaft, the first carrier can be fixed to the case and can be connected to the input shaft,
The multi-stage planetary gear train, wherein the second ring gear is connectable to the drive shaft. An input shaft;
An output shaft disposed parallel to the input shaft;
A counter shaft arranged parallel to the input shaft;
A drive shaft disposed parallel to the input shaft and the counter shaft;
A first planetary gear set and a second planetary gear set which are provided between the input shaft and the output shaft and convert the rotational speed of the input shaft into the rotational speed of the output shaft;
The first planetary gear set is disposed on the same axis as the counter shaft, and includes a first sun gear, a first ring gear, a first pinion engaged with the first ring gear and the first sun gear, A first carrier that pivotally supports the first pinion,
The second planetary gear set is disposed coaxially with the drive shaft, and a second sun gear, a second ring gear, a second pinion meshed with the second ring gear and the second sun gear, and the second A second carrier that pivotally supports the pinion,
A first gear pair connecting between the input shaft and the drive shaft;
A second gear pair connecting the input shaft and the counter shaft;
A first chain connecting between the first sun gear and the second sun gear;
A second chain connecting between the first ring gear and the second carrier;
The first sun gear and the second sun gear can be fixed to a case;
The second sun gear and the counter shaft are connectable,
One of the first ring gear and the second carrier is connected to the output shaft, and the first carrier can be fixed to the case and can be connected to the counter shaft.
The multi-stage planetary gear train, wherein the second ring gear is connectable to the drive shaft. The multi-stage planetary gear train according to claim 1 or 2, wherein the first sun gear and the input shaft or the counter shaft are connectable. An input shaft;
An output shaft disposed parallel to the input shaft;
A counter shaft arranged parallel to the input shaft;
A drive shaft disposed parallel to the input shaft and the counter shaft;
A first planetary gear set and a second planetary gear set which are provided between the input shaft and the output shaft and convert the rotational speed of the input shaft into the rotational speed of the output shaft;
The first planetary gear set is disposed coaxially with the counter shaft, and includes a first sun gear, a first ring gear, a first pinion meshed with the first ring gear and the first sun gear, and the first A first carrier that pivotally supports the pinion,
The second planetary gear set is disposed coaxially with the drive shaft, and a second sun gear, a second ring gear, a second outer pinion meshing with the second ring gear, the second outer pinion, and the second A second inner pinion meshed with the sun gear, and a second carrier that pivotally supports the second inner pinion and the second outer pinion,
A first gear pair connecting between the input shaft and the drive shaft;
A second gear pair connecting the input shaft and the counter shaft;
A first chain connecting between the first sun gear and the second carrier or the second sun gear;
A second chain connecting between the first carrier and the second ring gear;
The first sun gear and the second carrier or the second sun gear can be fixed to a case;
The second carrier and the drive shaft are connectable;
The first carrier and the second ring gear can be fixed to the case;
The first carrier and the counter shaft are connectable;
The first ring gear is connected to the output shaft;
The multi-stage planetary gear train, wherein the second sun gear is connectable to the drive shaft.

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