JP4333043B2

JP4333043B2 - Automatic transmission

Info

Publication number: JP4333043B2
Application number: JP2001060171A
Authority: JP
Inventors: 正宏早渕; 正明西田; 悟糟谷; 健次後藤; 敏彦青木
Original assignee: Aisin AW Co Ltd
Current assignee: Aisin AW Co Ltd
Priority date: 2001-03-05
Filing date: 2001-03-05
Publication date: 2009-09-16
Anticipated expiration: 2021-03-05
Also published as: JP2002266956A

Description

【０００１】
【発明の属する技術分野】
本発明は、入力軸に連結された減速用複式遊星歯車装置及び変速用複式遊星歯車装置の各要素に連結された制御クラッチ及び制御ブレーキを係脱して前記入力軸の回転を複数段に変速して出力軸に伝達する自動変速機に関する。
【０００２】
【従来の技術】
サンギヤ、該サンギヤに噛合するロングピニオンと該ロングピニオンに噛合する中間ピニオンとを支承するキャリヤ、前記ロングピニオンに噛合する第２リングギヤ及び前記中間ピニオンに噛合する第１リングギヤを有する減速用複式遊星歯車装置を設け、該減速用複式遊星歯車装置の第２リングギヤを入力軸に連結し、サンギヤ及び第１リングギヤを第１、第２制御ブレーキに夫々連結するとともに、第１制御クラッチにより相互に係脱可能に連結し、第１及び第２遊星歯車機構で変速用複式遊星歯車装置を構成し、第１遊星歯車機構のリングギヤと第２遊星歯車機構のキャリヤとを連結して入力軸に第２制御クラッチにより係脱可能に連結するとともに第３制御ブレーキに連結し、第２遊星歯車機構のリングギヤを減速用複式遊星歯車装置のキャリヤに連結し、第１及び第２遊星歯車機構の各サンギヤを第３制御クラッチで係脱可能に連結するとともに、第１遊星歯車機構のサンギヤを第４制御ブレーキに連結し、第１遊星歯車機構のキャリヤを出力軸に連結して前進７段、後退３段のギヤ比を成立する自動変速機が２０００−２６６１３８号公報に開示されている。
【０００３】
【発明が解決しようとする課題】
上記従来の自動変速機は、前進７段、後退３段の間でギヤ比を切り換えることが可能である。しかし、近年は燃費及び動力伝達性能向上を図るため、或いは運転者の嗜好にマッチしたギヤ比を得るために、適切に離間した前進８段のギヤ比を成立することができる自動変速機が求められている。
【０００４】
本発明は係る要望に応えるためになされたもので、適切に離間した前進８段のギヤ比を得ることができる自動変速機を提供することである。
【０００５】
【課題を解決するための手段】
上記の課題を解決するため、請求項１に記載の発明の構成上の特徴は、サンギヤ、該サンギヤに噛合するロングピニオンと該ロングピニオンに噛合する中間ピニオンとを支承するキャリヤ、前記ロングピニオンに噛合する第２リングギヤ及び前記中間ピニオンに噛合する第１リングギヤを有する減速用複式遊星歯車装置を設け、該減速用複式遊星歯車装置の第２リングギヤを入力軸に第４制御クラッチにより係脱可能に連結し、前記サンギヤ及び前記第１リングギヤを第１、第２制御ブレーキに夫々連結するとともに第１制御クラッチにより相互に係脱可能に連結し、第１及び第２遊星歯車機構で変速用複式遊星歯車装置を構成し、第１遊星歯車機構のリングギヤと第２遊星歯車機構のキャリヤとを連結して入力軸に第２制御クラッチにより係脱可能に連結するとともに第３制御ブレーキに連結し、第２遊星歯車機構のリングギヤを減速用複式遊星歯車装置のキャリヤに連結し、第１及び第２遊星歯車機構の各サンギヤを第３制御クラッチにより係脱可能に連結するとともに、第１遊星歯車機構のサンギヤを第４制御ブレーキに連結し、第１遊星歯車機構のキャリヤを出力軸に連結したことである。
【０００６】
請求項２に係る発明の構成上の特徴は、サンギヤ、該サンギヤに噛合するロングピニオンと該ロングピニオンに噛合する中間ピニオンとを支承するキャリヤ、前記ロングピニオンに噛合する第２リングギヤ及び前記中間ピニオンに噛合する第１リングギヤを有する減速用複式遊星歯車装置を設け、該減速用複式遊星歯車装置の第２リングギヤを入力軸に連結し、前記サンギヤ及び前記第１リングギヤを第１、第２制御ブレーキに夫々連結するとともに第１制御クラッチにより相互に係脱可能に連結し、第１及び第２遊星歯車機構で変速用複式遊星歯車装置を構成し、第１遊星歯車機構のリングギヤと第２遊星歯車機構のキャリヤとを連結して入力軸に第２制御クラッチにより係脱可能に連結するとともに第３制御ブレーキに連結し、第２遊星歯車機構のリングギヤを減速用複式遊星歯車装置のキャリヤに連結するとともに第５制御ブレーキに連結し、第１及び第２遊星歯車機構の各サンギヤを第３制御クラッチにより係脱可能に連結するとともに、第１遊星歯車機構のサンギヤを第４制御ブレーキに連結し、第１遊星歯車機構のキャリヤを出力軸に連結したことである。
【０００７】
【発明の作用・効果】
上記のように構成した請求項１に係る発明においては、サンギヤ、該サンギヤに噛合するロングピニオンと該ロングピニオンに噛合する中間ピニオンとを支承するキャリヤ、ロングピニオンに噛合する第２リングギヤ及び中間ピニオンに噛合する第１リングギヤを有する減速用複式遊星歯車装置の第２リングギヤを入力軸に第４制御クラッチで選択的に接続し、サンギヤ及び第１リングギヤを第１、第２制御ブレーキで選択的に回転規制するとともに第１制御クラッチで相互に選択的に接続し、変速用複式遊星歯車装置の第１遊星歯車機構のリングギヤと第２遊星歯車機構のキャリヤとを連結して入力軸に第２制御クラッチで選択的に接続するとともに第３制御ブレーキで選択的に回転規制し、第２遊星歯車機構のリングギヤを減速用複式遊星歯車装置のキャリヤに連結し、第１及び第２遊星歯車機構の各サンギヤを第３制御クラッチで選択的に接続し、第１遊星歯車機構のサンギヤの回転を第４制御ブレーキで選択的に規制し、第１遊星歯車機構のキャリヤを出力軸に連結したので、従来の自動変速機に１個の制御クラッチを加えるだけで、入力軸の回転を適切に離間した前進８段のギヤ比で変速して出力軸に伝達することができる自動変速機を提供することができる。さらに、高速段側のギヤ比を更に密にすることができるので、車速の高速度域でエンジン性能を最適に引き出すことができ、且つギヤチェンジ時のギヤ比の変化延いては出力トルクの変化が小さくなり、良好なフィーリングを得ることができる。
【０００８】
上記のように構成した請求項２に係る発明においては、サンギヤ、該サンギヤに噛合するロングピニオンと該ロングピニオンに噛合する中間ピニオンとを支承するキャリヤ、ロングピニオンに噛合する第２リングギヤ及び中間ピニオンに噛合する第１リングギヤを有する減速用複式遊星歯車装置の第２リングギヤを入力軸に連結し、サンギヤ及び第１リングギヤを第１、第２制御ブレーキで選択的に回転規制するとともに第１制御クラッチで相互に選択的に接続し、変速用複式遊星歯車装置の第１遊星歯車機構のリングギヤと第２遊星歯車機構のキャリヤとを連結して入力軸に第２制御クラッチで選択的に接続するとともに第３制御ブレーキで選択的に回転規制し、第２遊星歯車機構のリングギヤを減速用複式遊星歯車装置のキャリヤに連結するとともに第５制御ブレーキで選択的に回転規制し、第１及び第２遊星歯車機構の各サンギヤを第３制御クラッチで選択的に接続し、第１遊星歯車機構のサンギヤの回転を第４制御ブレーキで選択的に規制し、第１遊星歯車機構のキャリヤを出力軸に連結したので、従来の自動変速機に１個の制御ブレーキを加えるだけで、入力軸の回転を適切に離間した前進８段のギヤ比で変速して出力軸に伝達することができる。さらに、高速段側のギヤ比を更に密にすることができるので、車速の高速度域でエンジン性能を最適に引き出すことができ、且つギヤチェンジ時のギヤ比の変化延いては出力トルクの変化が小さくなり、良好なフィーリングを得ることができる。
【０００９】
【実施の形態】
以下、図面に基づいて本発明に係る自動変速機の第１の実施形態について説明する。図１において、１０は本発明に係る自動変速機で、例えば自動車のエンジンにより回転駆動される流体トルクコンバータ１１の出力回転を変速して駆動輪に伝達するために使用される。自動変速機１０は、車体に取り付けられたトランスミッションケース１２内に共通軸線１３上に順次支承された入力軸１５、減速用複式遊星歯車装置１６、変速用複式遊星歯車装置１７及び出力軸１８で構成されている。減速用複式遊星歯車装置１６は、ダブルピニオン型の遊星歯車機構２１及びシングルピニオン型の遊星歯車機構２２のサンギヤＳ１，Ｓ２及びキャリヤＣ１，Ｃ２を連結、共通化して構成されている。即ち、共通軸線１３上に回転可能に支承された共通のサンギヤＳ１，Ｓ２、サンギヤＳ１，Ｓ２と噛合するロングピニオン２３、このロングピニオン２３及びロングピニオン２３と噛合する中間ピニオン２４を回転可能に支承し共通軸線１３上に回転可能に支承された共通のキャリヤＣ１，Ｃ２、ロングピニオン２３及び中間ピニオン２４と夫々噛合し共通軸線１３上に回転可能に支承されたリングギヤＲ２，Ｒ１から構成されている。リングギヤＲ２は第４制御クラッチＣ−４により入力軸１５に係脱可能に連結されている。サンギヤＳ１，Ｓ２及び第１リングギヤＲ１は、ミッションケース１２に設けられた第１、第２制御ブレーキＢ−１，Ｂ−２に夫々連結されるとともに第１制御クラッチＣ−１により相互に係脱可能に連結されている。キャリヤＣ１，Ｃ２はワンウエイクラッチＦ−１を介してミッションケース１２に連結され、逆転を規制されている。
【００１０】
変速用複式遊星歯車装置１７は、共通軸線１３上に回転可能に支承されたサンギヤ、リングギヤ、サンギヤとリングギヤとに噛合するピニオンを支承するキャリヤで構成されるシングルピニオン型の第１及び第２遊星歯車機構２５，２６で構成されている。第１遊星歯車機構２５のリングギヤＲ３と第２遊星歯車機構２５のキャリヤＣ４とは連結されて連結要素Ｒ３，Ｃ４を形成し、連結要素Ｒ３，Ｃ４は入力軸１５に第２制御クラッチＣ−２により係脱可能に連結されるとともに、ミッションケース１２に設けられた第３制御ブレーキＢ−３に連結されている。第２遊星歯車機構２６のリングギヤＲ４は減速用複式遊星歯車装置１６のキャリヤＣ１，Ｃ２に連結されている。第１及び第２遊星歯車機構２５，２６の各サンギヤＳ３，Ｓ４は、第３制御クラッチＣ−３により相互に係脱可能に連結されている。第１遊星歯車機構２５のサンギヤＳ３は、ミッションケース１２に設けられた第４制御ブレーキＢ−４に連結されている。第１遊星歯車機構２１のキャリヤＣ３は出力軸１８に連結されている。
【００１１】
なお、流体トルクコンバータ１１のポンプインペラ３０は図略のエンジンによって回転駆動されてオイルを送り出し、ステータ３１がオイルの反力を受け止めてトルクをタービン３２に発生するようになっている。入力軸１５はタービン３２に連結されている。３３はポンプインペラ３０とタービン３２とを直結するロックアップクラッチである。
【００１２】
以上のように構成された自動変速機１０は、第１乃至第４制御クラッチＣ−１〜Ｃ−４を選択的に係脱し、第１乃至第４制御ブレーキＢ−１〜Ｂ−４を選択的に作動して減速用、変速用複式遊星歯車装置１６，１７の要素の回転を規制することにより、前進８段、後退３段のギヤ比を成立することができる。図２において、各変速段に対応する各制御クラッチ、制御ブレーキの欄に黒丸が付されている場合、制御クラッチであれば接続状態、制御ブレーキであれば回転規制状態にあることを示す。また、図２には、各変速段におけるギヤ比（入力軸１５の回転数／出力軸１８の回転数）及び変速段が１段アップしたときのギヤ比の増加割合（当変速段のギヤ比／前変速段のギヤ比）が示されている。
【００１３】
減速用複式遊星歯車装置１６のダブルピニオン型遊星歯車機構２１においては、サンギヤの回転数Ｎｓ、キャリヤの回転数Ｎｃ、リングギヤの回転数Ｎｒと遊星歯車機構のギヤ比λとの関係は、式（１）で示され、減速用複式遊星歯車装置１６のシングルピニオン型遊星歯車機構２２及びシングルピニオン型の第１、第２遊星歯車機構２５，２６においては、サンギヤの回転数Ｎｓ、キャリヤの回転数Ｎｃ、リングギヤの回転数Ｎｒと遊星歯車機構のギヤ比λとの関係は、式（２）で示され、各変速段におけるギヤ比は、式（１）、（２）に基づいて算出される。サンギヤＳ１，Ｓ２，Ｓ３，Ｓ４の歯数をＺｓ１，Ｚｓ２，Ｚｓ３，Ｚｓ４、リングギヤＲ１，Ｒ２，Ｒ３，Ｒ４の歯数をＺｒ１，Ｚｒ２，Ｚｒ３，Ｚｒ４とすると、減速用複式遊星歯車装置１６、変速用複式遊星歯車装置１７の遊星歯車機構２１，２２，２５，２６のギヤ比はλ１＝Ｚｓ１／Ｚｒ１，λ２＝Ｚｓ２／Ｚｒ２，λ３＝Ｚｓ３／Ｚｒ３，λ４＝Ｚｓ４／Ｚｒ４である。
Ｎｒ＝（１−λ）Ｎｃ＋λＮｓ・・・（１）
Ｎｒ＝（１＋λ）Ｎｃ−λＮｓ・・・（２）
【００１４】
第１乃至第４制御ブレーキＢ−１〜Ｂ−４を選択的に作動し、第１乃至第４制御クラッチＣ−１〜Ｃ−４を選択的に接続したとき、減速用複式遊星歯車装置１６及び変速用複式遊星歯車装置１７の各要素の速度比は、図３に示す速度線図のようになる。速度線図は、遊星歯車装置のサンギヤ、キャリヤ、リングギヤからなる各要素を横軸方向にギヤ比に対応させた間隔で配置し、縦軸方向に各要素に対応してその速度比を取ったものである。図３には、減速用及び変速用複式遊星歯車装置１６，１７の速度線図が左右に並べて記載されている。減速用複式遊星歯車装置１６では、サンギヤＳ１，Ｓ２、キャリヤＣ１，Ｃ２、がそれぞれ共通するので、Ｓ１，Ｓ２及びＣ１，Ｃ２がそれぞれ付された各１本の縦線上に共通のサンギヤＳ１，Ｓ２、共通のキャリヤＣ１，Ｃ２の速度比を表し、それぞれＲ１、Ｒ２が付された各１本の縦線上にリングギヤＲ１，Ｒ２の速度比を表す。ダブルピニオン型の遊星歯車機構２１については、キャリヤＣ１の縦線とサンギヤＳ１の縦線との間隔を１とみなし、リングギヤＲ１の縦線をキャリヤＣ１の縦線からサンギヤＳ１の縦線と同じ側に間隔λ１だけ離して配置する。シングルピニオン型の遊星歯車機構２２については、キャリヤＣ２の縦線とサンギヤＳ２の縦線との間隔を１とみなし、リングギヤＲ２の縦線をキャリヤＣ２の縦線からサンギヤＳ２の縦線の反対側に間隔λ２だけ離して配置する。
【００１５】
変速用複式遊星歯車装置１７では、第１、第２遊星歯車機構２５，２６のリングギヤＲ３とキャリヤＣ４とが連結され、サンギヤＳ３，Ｓ４が連結されるので、Ｒ３，Ｃ４及びＳ３，Ｓ４がそれぞれ付された各１本の縦線上に連結されたリングギヤＲ３とキャリヤＣ４、サンギヤＳ３，Ｓ４の速度比を表し、それぞれＲ４、Ｃ３が付された各１本の縦線上にリングギヤＲ４、キャリヤＣ３の速度比を表す。第１遊星歯車機構２５はシングルピニオン型であるので、リングギヤＲ３の縦線とサンギヤＳ３の縦線との間隔を１＋λ３とみなし、キャリヤＣ３の縦線をサンギヤＳ３の縦線からリングギヤＲ３の縦線の同じ側に間隔１だけ離して配置する。第２遊星歯車機構２６もシングルピニオン型であるので、キャリヤＣ４の縦線とサンギヤＳ４の縦線との間隔を１とみなし、リングギヤＲ４の縦線をキャリヤＣ４の縦線からサンギヤＳ４の縦線の反対側に間隔λ４だけ離して配置する。速度線図には、第１乃至第４制御ブレーキＢ−１〜Ｂ−４、第１乃至第４制御クラッチＣ−１〜Ｃ−４が選択的に作動された点にＢ−１〜Ｂ−４、Ｃ−１〜Ｃ−２が記入されている。
【００１６】
このように作成された変速用複式遊星歯車装置１７の速度線図において、４本の各縦線に対応する要素を縦線の並び順に第１、第２、第３、第４要素とする。第１実施形態の場合、第１要素としてのリングギヤＲ４は減速用複式遊星歯車装置１６の共通のキャリヤＣ１，Ｃ２に連結され、第２要素としての互いに連結されたリングギヤＲ３とキャリヤＣ４は第２制御クラッチＣ−２と第３制御ブレーキＢ−３とに並列に連結され、第３要素としてのキャリヤＣ３は出力軸１８に連結され、第４要素としてのサンギヤＳ３，Ｓ４は第３クラッチＣ−３により相互に係脱可能に連結され、サンギヤＳ３は第４制御ブレーキＢ−４に連結されている。
【００１７】
以下、各変速段の作動について説明する。前進第１変速段の場合、第２制御ブレーキＢ−２が作動されてリングギヤＲ１が回転規制され、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、減速用複式遊星歯車装置１６は、入力軸１５の回転より回転数の小さい第１減速回転をキャリヤＣ１，Ｃ２に生成する。この第１減速回転は第１要素であるリングギヤＲ４に伝達され、第３制御クラッチＣ−３が作動してサンギヤＳ３，Ｓ４が接続され、第４制御ブレーキＢ−４が作動してサンギヤＳ３，Ｓ４が回転規制されるので、第３要素であるキャリヤＣ３、出力軸１８は第１変速段のギヤ比５．４２８で正転駆動される。
【００１８】
前進第２変速段の場合、第１制御ブレーキＢ−１が作動されてサンギヤＳ１，Ｓ２が回転規制され、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、減速用複式遊星歯車装置１６は、回転数が入力軸１５より小さく第１減速回転より大きい第２減速回転をキャリヤＣ１，Ｃ２に生成する。この第２減速回転は第１要素であるリングギヤＲ４に伝達され、第３制御クラッチＣ−３が作動してサンギヤＳ３，Ｓ４が接続され、第４制御ブレーキＢ−４が作動してサンギヤＳ３，Ｓ４が回転規制されるので、第３要素であるキャリヤＣ３、出力軸１８は第２変速段のギヤ比３．０９２で正転駆動される。
【００１９】
前進第３変速段の場合、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、第１制御クラッチＣ−１が作動されてサンギヤＳ１，Ｓ２とリングギヤＲ１とが接続されるので、減速用複式遊星歯車装置１６は一体回転されて入力軸１５と同一回転数の入力回転をキャリヤＣ１，Ｃ２に生成する。この入力回転は第１要素であるリングギヤＲ４に伝達され、第３制御クラッチＣ−３が作動してサンギヤＳ３，Ｓ４が接続され、第４制御ブレーキＢ−４が作動してサンギヤＳ３，Ｓ４が回転規制されるので、第３要素であるキャリヤＣ３、出力軸１８は第３変速段のギヤ比２．０６１で正転駆動される。
【００２０】
前進第４変速段の場合、第２制御クラッチＣ−２が接続されてリングギヤＲ３が入力軸１５と同一回転数で回転され、第４制御ブレーキＢ−４が作動されてサンギヤＳ３が回転規制されるので、キャリヤＣ３、出力軸１８は第４変速段のギヤ比１．４０５で正転駆動される。このとき、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、第１制御クラッチＣ−１が作動されてサンギヤＳ１，Ｓ２とリングギヤＲ１とが接続されるので、減速用複式遊星歯車装置１６は入力回転で一体回転される。
【００２１】
前進第５変速段の場合、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、第１制御クラッチＣ−１が作動されてサンギヤＳ１，Ｓ２とリングギヤＲ１とが接続されるので、減速用複式遊星歯車装置１６は一体回転されて入力軸１５と同一回転数の入力回転をキャリヤＣ１，Ｃ２に生成する。この入力回転は第１要素であるリングギヤＲ４に伝達され、第２制御クラッチＣ−２が接続されて連結要素Ｒ３，Ｃ４が入力軸１５と同一回転数で回転され、第３制御クラッチＣ−３が作動してサンギヤＳ３，Ｓ４が接続されるので、変速用複式遊星歯車装置１７は一体回転され、キャリヤＣ３、出力軸１８は第５変速段のギヤ比１．０００で正転駆動される。
【００２２】
前進第６変速段の場合、第１制御ブレーキＢ−１が作動されてサンギヤＳ１，Ｓ２が回転規制され、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、減速用複式遊星歯車装置１６は、第２減速回転をキャリヤＣ１，Ｃ２に生成する。この第２減速回転は第１要素であるリングギヤＲ４に伝達され、連結要素Ｒ３，Ｃ４は第２制御クラッチＣ−２により入力軸１５に接続され、サンギヤＳ３，Ｓ４は第３制御クラッチＣ−３により接続されるので、キャリヤＣ３は入力回転と第２減速回転との回転差に応じて第６変速段のギヤ比０．８２９で正転駆動される。
【００２３】
前進第７変速段の場合、第２制御ブレーキＢ−２が作動されて第１リングギヤＲ１が回転規制され、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、減速用複式遊星歯車装置１６は、第１減速回転をキャリヤＣ１，Ｃ２に生成する。この第１減速回転は第１要素であるリングギヤＲ４に伝達され、連結要素Ｒ３，Ｃ４は第２制御クラッチＣ−２により入力軸１５に接続され、サンギヤＳ３，Ｓ４は第３制御クラッチＣ−３により接続されるので、キャリヤＣ３は入力回転と第１減速回転との回転差に応じて第７変速段のギヤ比０．７２３で正転駆動される。
【００２４】
前進第８変速段の場合、連結要素Ｒ３，Ｃ４は第２制御クラッチＣ−２により入力軸１５に接続されて入力回転され、第１制御クラッチＣ−１が接続されて一体化された減速用複式遊星歯車装置１６は第２制御ブレーキＢ−２の作動により回転規制されて第１要素であるリングギヤＲ４の回転を規制し、サンギヤＳ３，Ｓ４は第３制御クラッチＣ−３により接続されるので、キャリヤＣ３は第８変速段のギヤ比０．６１８で正転駆動される。
【００２５】
減速用複式遊星歯車装置１６を一体化して回転を規制することによりリングギヤＲ４の回転を規制するために、第１制御クラッチＣ−１を接続して第１制御ブレーキＢ−１を作動してもよく、又は第１及び第２制御ブレーキＢ−１，Ｂ−２を作動してもよい。
【００２６】
後退第１変速段の場合、第２制御ブレーキＢ−２が作動されて第１リングギヤＲ１が回転規制され、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、減速用複式遊星歯車装置１６は、第１減速回転をキャリヤＣ１，Ｃ２に生成する。この第１減速回転は第１要素であるリングギヤＲ４に伝達され、第３制御ブレーキＢ−３が作動して連結要素Ｒ３，Ｃ４の回転が規制され、サンギヤＳ３，Ｓ４は第３制御クラッチＣ−３で接続されるので、キャリヤＣ３と出力軸１８は後退第１変速段のギヤ比４．２６０で逆転駆動される。
【００２７】
後退第２変速段の場合、第１制御ブレーキＢ−１が作動されてサンギヤＳ１，Ｓ２が回転規制され、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、減速用複式遊星歯車装置１６は、第２減速回転をキャリヤＣ１，Ｃ２に生成する。この第２減速回転は第１要素であるリングギヤＲ４に伝達され、第３制御ブレーキＢ−３が作動して連結要素Ｒ３，Ｃ４の回転が規制され、サンギヤＳ３，Ｓ４は第３制御クラッチＣ−３で接続されるので、キャリヤＣ３と出力軸１８は後退第２変速段のギヤ比２．４２７で逆転駆動される。
【００２８】
後退第３変速段の場合、第４制御クラッチＣ−４が接続されて第２リングギヤＲ２が入力軸１５により回転駆動され、第１制御クラッチＣ−１が作動されてサンギヤＳ１，Ｓ２とリングギヤＲ１とが接続されるので、減速用複式遊星歯車装置１６は一体回転されて入力軸１５と同一回転数の入力回転をキャリヤＣ１，Ｃ２に生成する。この入力回転は第１要素であるリングギヤＲ４に伝達され、第３制御ブレーキＢ−３が作動して連結要素Ｒ３，Ｃ４の回転が規制され、サンギヤＳ３，Ｓ４は第３制御クラッチＣ−３で接続されるので、キャリヤＣ３と出力軸１８は後退第３変速段のギヤ比１．６１８で逆転駆動される。
【００２９】
入力軸１５の回転数を１とした場合の各変速段におけるサンギヤＳ１〜Ｓ４、キャリヤＣ１〜Ｃ４、及びリングギヤＲ１〜Ｒ４の回転比を示す図３の速度線図から明らかなように、各変速段における第３要素であるキャリヤＣ３の回転比すなわちギヤ比は、適当な間隔をもって配列し、本発明に係る自動変速機によれば適切に離間した前進８段、後退３段のギヤ比を得ることができる。さらに、変速段が１段アップしたときのギヤ比の増加割合は、図２に示すように、第１、第２変速段の間は１．７５６、第２、第３変速段の間は１．５００、第３、第４変速段の間は１．４６７、第４、第５変速段の間は１．４０５、第５、第６変速段の間は１．２０６、第６、第７変速段の間は１．１４７、第７、第８変速段の間は１．１７０となり、高速段ほど概ね増加割合が小さくなり、車速の高速度域でエンジン性能を最適に引き出すことができ、且つギヤチェンジ時のギヤ比の変化割合延いては出力トルクの変化割合が小さくなり、良好なフィーリングを得ることができる。また、最低変速段のギヤ比を最高変速段のギヤ比で割ったスプレッドは、８．７８と十分大きいので、低変速段において加速性能を高くし、高変速段において燃費をよくすることができる。
【００３０】
第１実施形態においては、減速用複式遊星歯車装置１６の第２リングギヤＲ２を第４制御クラッチＣ−４により入力軸１５に係脱可能に連結し、キャリヤＣ１，Ｃ２を変速用複式遊星歯車装置１７の第１要素であるリングギヤＲ４に連結するようにしているが、第２実施形態では、図４に示すように、第２リングギヤＲ２を入力軸１５に直接連結し、変速用複式遊星歯車装置１７の第１要素であるリングギヤＲ４をキャリヤＣ１，Ｃ２に連結するとともに第５制御ブレーキＢ−５に連結している。
【００３１】
第２実施形態における各変速段における各制御クラッチ、制御ブレーキの作動状態を図５に示す。各変速段におけるギヤ比、及び増加割合は第１実施形態の場合と同一である。速度線図は、図６に示すように、第１実施形態の場合と実質的に同じである。各変速段の作動は、前進第８変速段以外は、減速用複式遊星歯車装置１６の第２リングギヤＲ２が入力軸１５により第４制御クラッチＣ−４を介して回転駆動されるか、直接回転駆動されるかの相違点以外は同じであるので、説明を省略する。
【００３２】
前進第８変速段の場合、連結要素Ｒ３，Ｃ４は第２制御クラッチＣ−２により入力軸１５に接続されて入力回転され、第５制御ブレーキＢ−５の作動により第１要素であるリングギヤＲ４の回転が規制され、サンギヤＳ３，Ｓ４は第３制御クラッチＣ−３により接続されるので、キャリヤＣ３は第８変速段のギヤ比０．６１８で正転駆動される。
【図面の簡単な説明】
【図１】本発明に係る自動変速装置の第１実施形態を示すスケルトン図である。
【図２】第１実施形態の各速度段における制御ブレーキ及び制御クラッチの作動状態を示す図である。
【図３】第１実施形態の各速度段における遊星歯車装置の各要素の回転比を示す速度線図である。
【図４】第２実施形態を示すスケルトン図である。
【図５】第２実施形態の各速度段における制御ブレーキ及び制御クラッチの作動状態を示す図である。
【図６】第２実施形態の各速度段における遊星歯車装置の各要素の回転比を示す速度線図である。
【符号の説明】
１０・・・自動変速機、１１・・・流体トルクコンバータ、１２・・・トランスミッションケース、１３・・・共通軸線、１５・・・入力軸、１６・・・減速用複式遊星歯車装置、１７・・・変速用複式遊星歯車装置、１８・・・出力軸、２５・・・第１遊星歯車機構、２６・・・第２遊星歯車機構、Ｓ１，Ｓ２，Ｓ３，Ｓ４・・・サンギヤ、Ｃ１，Ｃ２，Ｃ３，Ｃ４・・・キャリヤ、Ｒ１，Ｒ２，Ｒ３，Ｒ４・・・リングギヤ、Ｒ３，Ｃ４・・・連結要素、Ｃ−１，Ｃ−２，Ｃ−３，Ｃ−４・・・第１、第２、第３、第４、制御クラッチ、Ｂ−１，Ｂ−２，Ｂ−３，Ｂ−４，Ｂ−５・・・第１、第２、第３、第４、第５制御ブレーキ。[0001]
BACKGROUND OF THE INVENTION
The present invention disengages a control clutch and a control brake connected to each element of a reduction type planetary gear unit and a speed change type planetary gear unit connected to an input shaft, and shifts the rotation of the input shaft to a plurality of stages. The present invention relates to an automatic transmission that transmits to an output shaft.
[0002]
[Prior art]
Reduction planetary gear having a sun gear, a carrier supporting a long pinion meshing with the sun gear and an intermediate pinion meshing with the long pinion, a second ring gear meshing with the long pinion, and a first ring gear meshing with the intermediate pinion And a second ring gear of the double planetary gear unit for speed reduction is connected to the input shaft, the sun gear and the first ring gear are connected to the first and second control brakes, respectively, and the first control clutch is engaged with and disengaged from each other. The first planetary gear mechanism and the second planetary gear mechanism constitute a shift type planetary gear device, and the ring gear of the first planetary gear mechanism and the carrier of the second planetary gear mechanism are coupled to perform second control on the input shaft. The second planetary gear mechanism is connected to the third control brake while being detachably connected by a clutch, and the double planetary gear unit for reduction is connected to the ring gear of the second planetary gear mechanism. Each of the sun gears of the first and second planetary gear mechanisms is detachably connected by the third control clutch, and the sun gear of the first planetary gear mechanism is connected to the fourth control brake, Japanese Patent Laid-Open No. 2000-266138 discloses an automatic transmission in which a gear mechanism carrier is connected to an output shaft to establish a forward gear ratio and a reverse gear ratio of three.
[0003]
[Problems to be solved by the invention]
The conventional automatic transmission can switch the gear ratio between 7 forward speeds and 3 reverse speeds. However, in recent years, there has been a demand for an automatic transmission that can establish a gear ratio of eight forwardly separated gears in order to improve fuel consumption and power transmission performance, or to obtain a gear ratio that matches the driver's preference. It has been.
[0004]
The present invention has been made in order to meet such a demand, and it is an object of the present invention to provide an automatic transmission capable of obtaining a gear ratio of 8 forwardly separated gears.
[0005]
[Means for Solving the Problems]
In order to solve the above-mentioned problem, the structural feature of the invention described in claim 1 is that a carrier that supports a sun gear, a long pinion meshing with the sun gear and an intermediate pinion meshing with the long pinion, and the long pinion A reduction type planetary gear unit having a second ring gear to be engaged and a first ring gear to be engaged with the intermediate pinion is provided, and the second ring gear of the reduction type planetary gear unit can be engaged and disengaged by a fourth control clutch with the input shaft as an input shaft. The sun gear and the first ring gear are connected to the first and second control brakes, and are connected to each other by a first control clutch so that they can be engaged and disengaged with each other. The gear device is configured, and the ring gear of the first planetary gear mechanism and the carrier of the second planetary gear mechanism are connected to each other by the second control clutch on the input shaft. The second planetary gear mechanism is connected to the third control brake, the ring gear of the second planetary gear mechanism is connected to the carrier of the double planetary gear unit for reduction, and the sun gears of the first and second planetary gear mechanisms are connected to the third control clutch. The sun gear of the first planetary gear mechanism is connected to the fourth control brake, and the carrier of the first planetary gear mechanism is connected to the output shaft.
[0006]
The structural features of the invention according to claim 2 are: a sun gear; a carrier that supports a long pinion that meshes with the sun gear and an intermediate pinion that meshes with the long pinion; a second ring gear that meshes with the long pinion; and the intermediate pinion A double planetary gear unit for reduction having a first ring gear meshing with the second ring gear is provided, the second ring gear of the double planetary gear unit for reduction is connected to an input shaft, and the sun gear and the first ring gear are connected to the first and second control brakes. And the first and second planetary gear mechanisms constitute a shift type planetary gear device, and the ring gear and the second planetary gear of the first planetary gear mechanism are connected to each other. The mechanism is connected to the carrier and connected to the input shaft by the second control clutch so as to be disengageable and connected to the third control brake. The ring gear of the vehicle mechanism is connected to the carrier of the double planetary gear unit for reduction and connected to the fifth control brake, and the sun gears of the first and second planetary gear mechanisms are detachably connected by the third control clutch, The sun gear of the first planetary gear mechanism is connected to the fourth control brake, and the carrier of the first planetary gear mechanism is connected to the output shaft.
[0007]
[Operation and effect of the invention]
In the invention according to claim 1 configured as described above, the sun gear, the carrier supporting the long pinion meshing with the sun gear and the intermediate pinion meshing with the long pinion, the second ring gear meshing with the long pinion, and the intermediate pinion The second ring gear of the double planetary gear unit for reduction having the first ring gear meshing with the input shaft is selectively connected to the input shaft by the fourth control clutch, and the sun gear and the first ring gear are selectively connected by the first and second control brakes. The rotation is restricted and the first control clutch selectively connects to each other, and the ring gear of the first planetary gear mechanism and the carrier of the second planetary gear mechanism of the double planetary gear device for speed change are connected to perform second control on the input shaft. The clutch is selectively connected and the third control brake selectively controls the rotation, and the ring gear of the second planetary gear mechanism is connected to the speed reduction compound idler. The sun gear of the first planetary gear mechanism is selectively connected by the third control clutch, and the rotation of the sun gear of the first planetary gear mechanism is selectively restricted by the fourth control brake. In addition, since the carrier of the first planetary gear mechanism is connected to the output shaft, the input shaft can be shifted at a gear ratio of 8 forward speeds with the rotation of the input shaft appropriately separated by adding only one control clutch to the conventional automatic transmission. Thus, an automatic transmission that can be transmitted to the output shaft can be provided. In addition, the gear ratio on the high speed side can be made denser, so that engine performance can be optimally achieved in the high speed range of the vehicle speed, and the change in the gear ratio at the time of the gear change and thus the change in the output torque. Becomes smaller and a good feeling can be obtained.
[0008]
In the invention according to claim 2 configured as described above, the sun gear, the carrier supporting the long pinion meshing with the sun gear and the intermediate pinion meshing with the long pinion, the second ring gear meshing with the long pinion, and the intermediate pinion The second ring gear of the reduction type planetary gear device having the first ring gear meshing with the first gear is connected to the input shaft, and the sun gear and the first ring gear are selectively restricted by the first and second control brakes and the first control clutch. And the ring gear of the first planetary gear mechanism and the carrier of the second planetary gear mechanism of the double planetary gear device for transmission are connected to each other and selectively connected to the input shaft by the second control clutch. The rotation is selectively restricted by the third control brake, and the ring gear of the second planetary gear mechanism is connected to the carrier of the double planetary gear unit for reduction. And the fifth control brake selectively restricts rotation, the sun gears of the first and second planetary gear mechanisms are selectively connected by the third control clutch, and the rotation of the sun gear of the first planetary gear mechanism is controlled by the fourth control. Since the carrier of the first planetary gear mechanism is connected to the output shaft by selectively restricting with the brake, the forward movement 8 in which the rotation of the input shaft is appropriately separated by adding only one control brake to the conventional automatic transmission. The gear can be shifted with the gear ratio and transmitted to the output shaft. In addition, the gear ratio on the high speed side can be made denser, so that engine performance can be optimally achieved in the high speed range of the vehicle speed, and the change in the gear ratio at the time of the gear change and thus the change in the output torque. Becomes smaller and a good feeling can be obtained.
[0009]
Embodiment
Hereinafter, a first embodiment of an automatic transmission according to the present invention will be described with reference to the drawings. In FIG. 1, reference numeral 10 denotes an automatic transmission according to the present invention, which is used, for example, for shifting the output rotation of a fluid torque converter 11 driven to rotate by an automobile engine and transmitting it to drive wheels. The automatic transmission 10 includes an input shaft 15 that is sequentially supported on a common axis 13 in a transmission case 12 attached to a vehicle body, a double planetary gear device 16 for reduction, a double planetary gear device 17 for transmission, and an output shaft 18. Has been. The reduction type double planetary gear unit 16 is configured by connecting and sharing the sun gears S1 and S2 and the carriers C1 and C2 of the double pinion type planetary gear mechanism 21 and the single pinion type planetary gear mechanism 22. That is, the common sun gears S1 and S2 rotatably supported on the common axis 13, the long pinion 23 meshing with the sun gears S1 and S2, and the long pinion 23 and the intermediate pinion 24 meshing with the long pinion 23 are rotatably supported. The ring gears R2 and R1 are engaged with the common carriers C1 and C2, which are rotatably supported on the common axis 13, and the long pinion 23 and the intermediate pinion 24, respectively, and are rotatably supported on the common axis 13. . The ring gear R2 is detachably connected to the input shaft 15 by a fourth control clutch C-4. The sun gears S1 and S2 and the first ring gear R1 are connected to first and second control brakes B-1 and B-2 provided in the transmission case 12, respectively, and are engaged and disengaged with each other by the first control clutch C-1. Connected as possible. The carriers C1 and C2 are connected to the transmission case 12 via the one-way clutch F-1 and are prevented from rotating in reverse.
[0010]
The shift type planetary gear unit 17 is a single pinion type first and second planetary planetary gear comprising a sun gear rotatably supported on a common axis 13, a ring gear, and a carrier supporting a pinion meshing with the sun gear and the ring gear. It consists of gear mechanisms 25 and 26. The ring gear R3 of the first planetary gear mechanism 25 and the carrier C4 of the second planetary gear mechanism 25 are connected to form connecting elements R3 and C4. The connecting elements R3 and C4 are connected to the input shaft 15 with the second control clutch C-2. And is connected to a third control brake B-3 provided in the transmission case 12. The ring gear R4 of the second planetary gear mechanism 26 is connected to the carriers C1 and C2 of the double planetary gear unit 16 for reduction. The sun gears S3 and S4 of the first and second planetary gear mechanisms 25 and 26 are detachably connected to each other by a third control clutch C-3. The sun gear S3 of the first planetary gear mechanism 25 is connected to a fourth control brake B-4 provided in the mission case 12. The carrier C3 of the first planetary gear mechanism 21 is connected to the output shaft 18.
[0011]
The pump impeller 30 of the fluid torque converter 11 is rotationally driven by an unillustrated engine to send out oil, and the stator 31 receives the reaction force of the oil and generates torque in the turbine 32. The input shaft 15 is connected to the turbine 32. Reference numeral 33 denotes a lock-up clutch that directly connects the pump impeller 30 and the turbine 32.
[0012]
The automatic transmission 10 configured as described above selectively disengages the first to fourth control clutches C-1 to C-4 and selects the first to fourth control brakes B-1 to B-4. The gear ratio of eight forward speeds and three reverse speeds can be established by restricting the rotation of the elements of the double planetary gear units 16 and 17 for speed reduction and speed change. In FIG. 2, when each control clutch and control brake column corresponding to each gear stage is marked with a black circle, this indicates that the control clutch is in the connected state and the control brake is in the rotation restricted state. FIG. 2 also shows the gear ratio at each gear stage (the rotational speed of the input shaft 15 / the rotational speed of the output shaft 18) and the rate of increase in the gear ratio when the gear stage is increased by one (the gear ratio of the current gear stage). / Gear ratio of the front gear).
[0013]
In the double pinion type planetary gear mechanism 21 of the double planetary gear unit 16 for reduction, the relationship among the rotational speed Ns of the sun gear, the rotational speed Nc of the carrier, the rotational speed Nr of the ring gear and the gear ratio λ of the planetary gear mechanism is 1), the single-pinion type planetary gear mechanism 22 and the single-pinion type first and second planetary gear mechanisms 25 and 26 of the double planetary gear unit 16 for reduction use a sun gear rotation speed Ns and a carrier rotation speed. The relationship between Nc, the rotational speed Nr of the ring gear and the gear ratio λ of the planetary gear mechanism is expressed by Expression (2), and the gear ratio at each gear stage is calculated based on Expressions (1) and (2). . If the number of teeth of the sun gears S1, S2, S3, S4 is Zs1, Zs2, Zs3, Zs4, and the number of teeth of the ring gears R1, R2, R3, R4 is Zr1, Zr2, Zr3, Zr4, the double planetary gear unit 16 for reduction, The gear ratios of the planetary gear mechanisms 21, 22, 25, and 26 of the shift type planetary gear unit 17 are λ1 = Zs1 / Zr1, λ2 = Zs2 / Zr2, λ3 = Zs3 / Zr3, and λ4 = Zs4 / Zr4.
Nr = (1-λ) Nc + λNs (1)
Nr = (1 + λ) Nc−λNs (2)
[0014]
When the first to fourth control brakes B-1 to B-4 are selectively operated and the first to fourth control clutches C-1 to C-4 are selectively connected, the double planetary gear unit 16 for speed reduction is used. And the speed ratio of each element of the compound planetary gear unit 17 for speed change is as shown in the speed diagram shown in FIG. In the speed diagram, the elements including the sun gear, the carrier, and the ring gear of the planetary gear device are arranged at intervals corresponding to the gear ratio in the horizontal axis direction, and the speed ratio is calculated in correspondence with each element in the vertical axis direction. Is. In FIG. 3, the speed diagrams of the double planetary gear units 16 and 17 for reduction and transmission are shown side by side. In the double planetary gear unit 16 for reduction, the sun gears S1, S2 and the carriers C1, C2 are common to each other. Therefore, the common sun gears S1, S2 are arranged on one vertical line to which S1, S2 and C1, C2 are respectively attached. Represents the speed ratio of the common carriers C1 and C2, and represents the speed ratio of the ring gears R1 and R2 on each one vertical line with R1 and R2 respectively. Regarding the double pinion type planetary gear mechanism 21, the distance between the vertical line of the carrier C1 and the vertical line of the sun gear S1 is regarded as 1, and the vertical line of the ring gear R1 is the same side as the vertical line of the sun gear S1 from the vertical line of the carrier C1. Are spaced apart by a distance λ1. For the single pinion type planetary gear mechanism 22, the distance between the vertical line of the carrier C2 and the vertical line of the sun gear S2 is regarded as 1, and the vertical line of the ring gear R2 is opposite to the vertical line of the sun gear S2 from the vertical line of the carrier C2. Are spaced apart by a distance λ2.
[0015]
In the shift type planetary gear unit 17, the ring gear R3 and the carrier C4 of the first and second planetary gear mechanisms 25 and 26 are connected and the sun gears S3 and S4 are connected, so that R3, C4 and S3, S4 are respectively connected. The speed ratio of the ring gear R3, the carrier C4, and the sun gears S3 and S4 connected on each one vertical line attached is shown, and the ring gear R4 and the carrier C3 on each one vertical line attached with R4 and C3, respectively. Represents speed ratio. Since the first planetary gear mechanism 25 is a single pinion type, the distance between the vertical line of the ring gear R3 and the vertical line of the sun gear S3 is regarded as 1 + λ3, and the vertical line of the carrier C3 is changed from the vertical line of the sun gear S3 to the vertical line of the ring gear R3. Are spaced apart on the same side by a distance of 1. Since the second planetary gear mechanism 26 is also a single pinion type, the distance between the vertical line of the carrier C4 and the vertical line of the sun gear S4 is regarded as 1, and the vertical line of the ring gear R4 is changed from the vertical line of the carrier C4 to the vertical line of the sun gear S4. Are spaced apart from each other by a distance λ4. In the speed diagram, B-1 to B- are the points where the first to fourth control brakes B-1 to B-4 and the first to fourth control clutches C-1 to C-4 are selectively operated. 4, C-1 to C-2 are entered.
[0016]
In the velocity diagram of the shift type planetary gear unit 17 created in this way, the elements corresponding to the four vertical lines are defined as the first, second, third, and fourth elements in the order of the vertical lines. In the case of the first embodiment, the ring gear R4 as the first element is connected to the common carriers C1 and C2 of the double planetary gear unit 16 for reduction, and the ring gear R3 and the carrier C4 connected to each other as the second element are the second elements. The control clutch C-2 and the third control brake B-3 are connected in parallel, the carrier C3 as the third element is connected to the output shaft 18, and the sun gears S3 and S4 as the fourth element are connected to the third clutch C- 3, the sun gear S3 is connected to the fourth control brake B-4.
[0017]
Hereinafter, the operation of each gear stage will be described. In the case of the first forward speed, the second control brake B-2 is operated to restrict the rotation of the ring gear R1, the fourth control clutch C-4 is connected, and the second ring gear R2 is rotationally driven by the input shaft 15. The reduction type planetary gear unit 16 generates a first reduction rotation having a rotation speed smaller than that of the input shaft 15 on the carriers C1 and C2. The first reduced speed rotation is transmitted to the ring gear R4 as the first element, the third control clutch C-3 is operated to connect the sun gears S3 and S4, and the fourth control brake B-4 is operated to operate the sun gear S3, S3. Since the rotation of S4 is restricted, the carrier C3, which is the third element, and the output shaft 18 are driven to rotate forward at a gear ratio of 5.428 of the first gear.
[0018]
In the case of the second forward speed, the first control brake B-1 is actuated to restrict the rotation of the sun gears S1, S2, the fourth control clutch C-4 is connected, and the second ring gear R2 is driven to rotate by the input shaft 15. Then, the double planetary gear unit 16 for speed reduction generates the second reduced speed rotation on the carriers C1 and C2 whose rotational speed is smaller than that of the input shaft 15 and larger than the first reduced speed rotation. This second reduced speed rotation is transmitted to the ring gear R4, which is the first element, the third control clutch C-3 is actuated to connect the sun gears S3 and S4, and the fourth control brake B-4 is actuated to act on the sun gear S3, Since the rotation of S4 is restricted, the carrier C3, which is the third element, and the output shaft 18 are driven to rotate forward at a gear ratio of 3.092 of the second gear.
[0019]
In the case of the third forward speed, the fourth control clutch C-4 is connected, the second ring gear R2 is rotationally driven by the input shaft 15, and the first control clutch C-1 is operated to cause the sun gears S1, S2 and the ring gear R1 to rotate. Are connected to each other, and the reduction type planetary gear unit 16 is integrally rotated to generate an input rotation having the same rotational speed as that of the input shaft 15 on the carriers C1 and C2. This input rotation is transmitted to the ring gear R4 which is the first element, the third control clutch C-3 is operated and the sun gears S3 and S4 are connected, and the fourth control brake B-4 is operated and the sun gears S3 and S4 are connected. Since the rotation is restricted, the carrier C3, which is the third element, and the output shaft 18 are driven to rotate forward at a gear ratio 2.061 of the third gear.
[0020]
In the case of the fourth forward speed, the second control clutch C-2 is connected, the ring gear R3 is rotated at the same rotational speed as the input shaft 15, the fourth control brake B-4 is operated, and the sun gear S3 is restricted from rotating. Therefore, the carrier C3 and the output shaft 18 are driven to rotate forward at a gear ratio of 1.405 at the fourth speed. At this time, the fourth control clutch C-4 is connected and the second ring gear R2 is rotationally driven by the input shaft 15, and the first control clutch C-1 is operated to connect the sun gears S1 and S2 and the ring gear R1. Therefore, the double planetary gear unit 16 for reduction is integrally rotated by the input rotation.
[0021]
In the case of the fifth forward speed, the fourth control clutch C-4 is connected and the second ring gear R2 is rotationally driven by the input shaft 15, and the first control clutch C-1 is operated to cause the sun gears S1 and S2 and the ring gear R1 to rotate. Are connected to each other, and the reduction type planetary gear unit 16 is integrally rotated to generate an input rotation having the same rotational speed as that of the input shaft 15 on the carriers C1 and C2. This input rotation is transmitted to the ring gear R4, which is the first element, the second control clutch C-2 is connected, the coupling elements R3, C4 are rotated at the same rotational speed as the input shaft 15, and the third control clutch C-3 Is operated and the sun gears S3 and S4 are connected, so that the shift type planetary gear unit 17 is rotated integrally, and the carrier C3 and the output shaft 18 are driven to rotate forward at a gear ratio of 1.000 of the fifth gear.
[0022]
In the case of the sixth forward speed, the first control brake B-1 is operated to restrict the rotation of the sun gears S1 and S2, the fourth control clutch C-4 is connected, and the second ring gear R2 is driven to rotate by the input shaft 15. Then, the double planetary gear unit 16 for speed reduction generates the second reduced speed rotation on the carriers C1 and C2. This second reduced speed rotation is transmitted to the ring gear R4 as the first element, the connecting elements R3 and C4 are connected to the input shaft 15 by the second control clutch C-2, and the sun gears S3 and S4 are connected to the third control clutch C-3. Therefore, the carrier C3 is driven to rotate forward at a gear ratio of 0.829 of the sixth gear according to the rotation difference between the input rotation and the second reduced rotation.
[0023]
In the case of the seventh forward speed, the second control brake B-2 is operated to restrict the rotation of the first ring gear R1, the fourth control clutch C-4 is connected, and the second ring gear R2 is driven to rotate by the input shaft 15. Then, the double planetary gear unit 16 for reduction generates the first reduction rotation on the carriers C1 and C2. This first reduced speed rotation is transmitted to the ring gear R4 as the first element, the connecting elements R3 and C4 are connected to the input shaft 15 by the second control clutch C-2, and the sun gears S3 and S4 are connected to the third control clutch C-3. Therefore, the carrier C3 is driven to rotate forward at a gear ratio of 0.723 of the seventh shift stage according to the rotation difference between the input rotation and the first reduction rotation.
[0024]
In the case of the eighth forward shift speed, the coupling elements R3 and C4 are connected to the input shaft 15 by the second control clutch C-2 for input rotation, and the first control clutch C-1 is connected and integrated for deceleration. The double planetary gear unit 16 is restricted in rotation by the operation of the second control brake B-2 and restricts the rotation of the ring gear R4 as the first element, and the sun gears S3 and S4 are connected by the third control clutch C-3. The carrier C3 is driven to rotate forward at a gear ratio 0.618 of the eighth gear.
[0025]
Even if the first control brake C-1 is operated and the first control brake B-1 is operated in order to restrict the rotation of the ring gear R4 by integrating the reduction type planetary gear unit 16 and restricting the rotation. Alternatively, the first and second control brakes B-1 and B-2 may be operated.
[0026]
In the case of the reverse first shift speed, the second control brake B-2 is operated to restrict the rotation of the first ring gear R1, the fourth control clutch C-4 is connected, and the second ring gear R2 is driven to rotate by the input shaft 15. Then, the double planetary gear unit 16 for reduction generates the first reduction rotation on the carriers C1 and C2. This first decelerated rotation is transmitted to the first element ring gear R4, the third control brake B-3 is actuated to restrict the rotation of the coupling elements R3, C4, and the sun gears S3, S4 are connected to the third control clutch C--. 3, the carrier C <b> 3 and the output shaft 18 are driven in reverse at a gear ratio of 4.260 of the reverse first gear.
[0027]
In the case of the second reverse speed, the first control brake B-1 is operated to restrict the rotation of the sun gears S1 and S2, the fourth control clutch C-4 is connected, and the second ring gear R2 is driven to rotate by the input shaft 15. Then, the double planetary gear unit 16 for speed reduction generates the second reduced speed rotation on the carriers C1 and C2. The second decelerated rotation is transmitted to the ring gear R4 as the first element, the third control brake B-3 is actuated to restrict the rotation of the coupling elements R3 and C4, and the sun gears S3 and S4 are connected to the third control clutch C- 3, the carrier C3 and the output shaft 18 are driven to rotate in reverse at a gear ratio of 2.427 for the second reverse speed.
[0028]
In the case of the reverse third shift speed, the fourth control clutch C-4 is connected, the second ring gear R2 is rotationally driven by the input shaft 15, the first control clutch C-1 is operated, and the sun gears S1, S2 and the ring gear R1 are operated. Are connected to each other, and the reduction type planetary gear unit 16 is integrally rotated to generate an input rotation having the same rotational speed as that of the input shaft 15 on the carriers C1 and C2. This input rotation is transmitted to the ring gear R4 as the first element, the third control brake B-3 is actuated to restrict the rotation of the coupling elements R3 and C4, and the sun gears S3 and S4 are driven by the third control clutch C-3. Since they are connected, the carrier C3 and the output shaft 18 are driven to rotate in reverse at the reverse third gear stage gear ratio of 1.618.
[0029]
As is apparent from the speed diagram of FIG. 3 showing the rotation ratios of the sun gears S1 to S4, the carriers C1 to C4, and the ring gears R1 to R4 at each shift speed when the rotational speed of the input shaft 15 is 1. The rotation ratio, that is, the gear ratio of the carrier C3, which is the third element in the stage, is arranged at an appropriate interval, and according to the automatic transmission according to the present invention, the gear ratio of the eight forward stages and the three reverse stages are obtained. be able to. Further, as shown in FIG. 2, the increase ratio of the gear ratio when the gear stage is increased by one stage is 1.756 between the first and second gear stages, and 1 between the second and third gear stages. .500, 1.467 between the third and fourth gears, 1.405 between the fourth and fifth gears, 1.206, sixth, seventh between the fifth and sixth gears. 1.147 between the gears and 1.170 between the seventh and eighth gears, the rate of increase is generally smaller at higher speeds, and the engine performance can be optimally extracted in the high speed range of the vehicle speed. Further, the change ratio of the gear ratio at the time of the gear change is extended, and the change ratio of the output torque becomes small, and a good feeling can be obtained. Further, the spread obtained by dividing the gear ratio of the lowest gear stage by the gear ratio of the highest gear stage is sufficiently large as 8.78, so that the acceleration performance can be improved at the low gear stage and the fuel efficiency can be improved at the high gear stage. .
[0030]
In the first embodiment, the second ring gear R2 of the reduction type planetary gear unit 16 is detachably connected to the input shaft 15 by the fourth control clutch C-4, and the carriers C1 and C2 are connected to the transmission type planetary gear unit. In the second embodiment, as shown in FIG. 4, the second ring gear R2 is directly connected to the input shaft 15, and a double planetary gear device for transmission is connected. A ring gear R4 as a first element 17 is connected to the carriers C1 and C2 and to the fifth control brake B-5.
[0031]
FIG. 5 shows operating states of the control clutches and the control brakes at the respective shift speeds in the second embodiment. The gear ratio and the increase rate at each shift stage are the same as those in the first embodiment. The velocity diagram is substantially the same as in the first embodiment, as shown in FIG. As for the operation of each gear stage, except for the forward eighth gear stage, the second ring gear R2 of the double planetary gear unit 16 for deceleration is driven to rotate by the input shaft 15 via the fourth control clutch C-4 or directly rotated. Since it is the same except for the difference in whether it is driven, the description is omitted.
[0032]
In the case of the eighth forward shift speed, the coupling elements R3 and C4 are connected to the input shaft 15 by the second control clutch C-2 to be input and rotated, and the ring gear R4 that is the first element is actuated by the fifth control brake B-5. , And the sun gears S3 and S4 are connected by the third control clutch C-3, so that the carrier C3 is driven forward at a gear ratio 0.618 of the eighth gear.
[Brief description of the drawings]
FIG. 1 is a skeleton diagram showing a first embodiment of an automatic transmission according to the present invention.
FIG. 2 is a diagram illustrating operating states of a control brake and a control clutch at each speed stage according to the first embodiment.
FIG. 3 is a velocity diagram showing a rotation ratio of each element of the planetary gear device at each speed stage of the first embodiment.
FIG. 4 is a skeleton diagram showing a second embodiment.
FIG. 5 is a diagram illustrating operating states of a control brake and a control clutch at each speed stage according to the second embodiment.
FIG. 6 is a velocity diagram showing a rotation ratio of each element of the planetary gear device at each speed stage of the second embodiment.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 10 ... Automatic transmission, 11 ... Fluid torque converter, 12 ... Transmission case, 13 ... Common axis, 15 ... Input shaft, 16 ... Double-type planetary gear unit for deceleration, 17. ..Double planetary gear unit for speed change, 18 ... output shaft, 25 ... first planetary gear mechanism, 26 ... second planetary gear mechanism, S1, S2, S3, S4 ... sun gear, C1, C2, C3, C4 ... carrier, R1, R2, R3, R4 ... ring gear, R3, C4 ... coupling element, C-1, C-2, C-3, C-4 ... No. 1st, 2nd, 3rd, 4th, control clutch, B-1, B-2, B-3, B-4, B-5 ... 1st, 2nd, 3rd, 4th, 5th Control brake.

Claims

Reduction planetary gear having a sun gear, a carrier supporting a long pinion meshing with the sun gear and an intermediate pinion meshing with the long pinion, a second ring gear meshing with the long pinion, and a first ring gear meshing with the intermediate pinion A second ring gear of the double planetary gear unit for reduction is connected to the input shaft by a fourth control clutch so as to be disengageable, and the sun gear and the first ring gear are connected to the first and second control brakes, respectively. In addition, the first and second planetary gear mechanisms are coupled to each other so as to be disengageable with each other, and a shift type planetary gear device is constituted by the first and second planetary gear mechanisms, and the ring gear of the first planetary gear mechanism and the carrier of the second planetary gear mechanism; And the second planetary gear connected to the input shaft and releasably connected by the second control clutch and to the third control brake. The ring gear of the structure is connected to the carrier of the double planetary gear unit for reduction, the sun gears of the first and second planetary gear mechanisms are detachably connected by the third control clutch, and the sun gear of the first planetary gear mechanism is connected to the first gear mechanism. 4. An automatic transmission connected to a control brake, wherein the carrier of the first planetary gear mechanism is connected to an output shaft.

Reduction planetary gear having a sun gear, a carrier supporting a long pinion meshing with the sun gear and an intermediate pinion meshing with the long pinion, a second ring gear meshing with the long pinion, and a first ring gear meshing with the intermediate pinion And a second ring gear of the double planetary gear unit for speed reduction is connected to the input shaft, the sun gear and the first ring gear are connected to the first and second control brakes, respectively, and the first control clutch engages with each other. The first and second planetary gear mechanisms are connected in a detachable manner to form a shift type planetary gear device, and the ring gear of the first planetary gear mechanism and the carrier of the second planetary gear mechanism are connected to the second input shaft. The control clutch is detachably connected to the third control brake, and the ring planetary gear mechanism is connected to the double planetary gear for reduction. The sun gear of the first planetary gear mechanism is connected to the carrier of the vehicle device and the fifth control brake, the sun gears of the first and second planetary gear mechanisms are detachably connected by the third control clutch, and the sun gear of the first planetary gear mechanism is connected to the second gear. 4. An automatic transmission connected to a control brake, wherein the carrier of the first planetary gear mechanism is connected to an output shaft.