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JP4590742B2 - Automatic transmission - Google Patents

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Publication number
JP4590742B2
JP4590742B2 JP2001011855A JP2001011855A JP4590742B2 JP 4590742 B2 JP4590742 B2 JP 4590742B2 JP 2001011855 A JP2001011855 A JP 2001011855A JP 2001011855 A JP2001011855 A JP 2001011855A JP 4590742 B2 JP4590742 B2 JP 4590742B2
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rotation
gear
planetary gear
stage
ring gear
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JP2002213545A (en
Inventor
正宏 早渕
正明 西田
悟 糟谷
健次 後藤
敏彦 青木
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Aisin AW Co Ltd
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Aisin AW Co Ltd
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H2200/00—Transmissions for multiple ratios
    • F16H2200/003—Transmissions for multiple ratios characterised by the number of forward speeds
    • F16H2200/006—Transmissions for multiple ratios characterised by the number of forward speeds the gear ratios comprising eight forward speeds

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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】
互いに噛合する一対のピニオンを支承するキャリヤ、該一対のピニオンの一方と噛合するサンギヤ、他方と噛合するリングギヤを有するダブルピニオン型の減速用遊星歯車装置と、前段及び後段サンギヤ、前段及び後段リングギヤ、前段サンギヤと前段リングギヤとに噛合するピニオンを支承する前段キャリヤ、後段サンギヤと後段リングギヤとに噛合するピニオンを支承する後段キャリヤを有し、前段及び後段サンギヤを互いに連結し、前段キャリヤと後段リングギヤとを連結した変速用複式遊星歯車装置を共通軸線上に設け、後段キャリヤに出力軸を連結し、入力軸の回転を互いに連結された前段及び後段サンギヤ、互いに連結された前段キャリヤ及び後段リングギヤに第１、第２制御クラッチにより選択的に伝達し、減速用遊星歯車装置のリングギヤの回転を前段リングギヤに伝達し、減速用遊星歯車装置のキャリヤと、リングギヤと、変速用複式遊星歯車装置の互いに連結された前段キャリヤ及び後段リングギヤとの回転を第１、第２、第３制御ブレーキにより夫々選択的に規制して前進５段、後退１段のギヤ比を成立する自動変速機が特開平２−１２９４４６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５段、後退１段の間でギヤ比を円滑に安定して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６段以上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６段以上、後退１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、入力軸と、該入力軸に連結され該入力軸の回転より回転数が小さい第１回転及び該第１回転より回転数が小さい第２回転を生成する歯車減速装置と、速度線図においてギヤ比に対応した間隔で順次並べられた４個の要素に並び順にそれぞれ対応する第１、第２、第３及び第４要素を有する変速用複式遊星歯車装置と、前記入力軸の回転を第２、第４要素に選択的に伝達する第２、第１制御クラッチと、前記第１及び第２回転の一方を選択して前記第１要素に伝達する回転選択手段と、前記第１、第２要素の回転を選択的に規制する第１及び第２制御ブレーキと、前記第３要素に連結された出力軸とを備え、前記変速用複式遊星歯車装置は、前記歯車減速装置から前記第１回転及び前記第２回転の一方が前記回転選択手段によって選択されて前記第１要素に伝達された場合に、前記入力軸の回転が前記第１制御クラッチで前記第４要素に伝達されることにより前記入力軸の回転数より前記出力軸の回転数が小さい減速段を形成し、前記歯車減速装置から前記第１回転及び前記第２回転の一方が前記回転選択手段によって選択されて前記第１要素に伝達された場合に、前記入力軸の回転が前記第２制御クラッチで前記第２要素に伝達されることにより前記入力軸の回転数より前記出力軸の回転数が大きい増速段を形成することである。
【０００６】
請求項２に係る発明の構成上の特徴は、請求項１に記載の自動変速機において、入力軸に連結されたサンギヤ、該サンギヤと噛合する小径ピニオン及び大径ピニオンからなる段付ピニオンを支承するキャリヤ、小径ピニオンと噛合する小径リングギヤ並びに前記大径ピニオンと噛合する大径リングギヤからなる減速用複式遊星歯車装置で前記歯車減速装置を構成し、該減速用複式遊星歯車装置のキャリヤを前記変速用複式遊星歯車装置の第１要素に連結し、前記小径及び大径リングギヤの回転を夫々選択的に規制して前記キャリヤに前記第１、第２回転を選択的に生成させる第１、第２回転生成制御ブレーキで前記回転選択手段を構成したことである。
【０００７】
請求項３に係る発明の構成上の特徴は、請求項１に記載の自動変速機において、入力軸に連結されたサンギヤ、該サンギヤと噛合するロングピニオン及び該ロングピニオンと噛合する中間ピニオンを支承するキャリヤ、前記ロングピニオンと噛合する第１リングギヤ並びに前記中間ピニオンと噛合する第２リングギヤからなる減速用複式遊星歯車装置で前記歯車減速装置を構成し、該減速用複式遊星歯車装置の第２リングギヤを前記変速用複式遊星歯車装置の第１要素に連結し、前記第１リングギヤ及びキャリヤの回転を夫々選択的に規制して前記第２リングギヤに前記第１、第２回転を選択的に生成させる第１、第２回転生成制御ブレーキで前記回転選択手段を構成したことである。
【０００８】
請求項４に係る発明の構成上の特徴は、請求項１に記載の自動変速機において、入力軸に固定された大径、中径、小径歯車、前記変速用遊星歯車装置と同心に回転可能に支承され前記大径、中径、小径歯車と夫々噛合して入力回転、該入力回転より回転数が小さい第１回転、該第１回転より回転数が小さい第２回転を夫々生成する入力歯車、第１歯車、第２歯車からなる減速用歯車列で前記歯車減速装置を構成し、前記入力歯車を前記第２、第１制御クラッチに連結して前記入力回転を前記変速用複式遊星歯車装置の第２、第４要素に選択的に伝達し、前記第１、第２歯車を前記第１要素に夫々係脱可能に連結する第１、第２回転伝達制御クラッチにより前記第１及び第２回転の一方を選択して前記第１要素に伝達する回転選択手段を構成したことである。
【０００９】
請求項５に係る発明の構成上の特徴は、請求項１乃至請求項４のいずれかに記載の自動変速機において、前記変速用複式遊星歯車装置を前後段２組のシングルピニオン型の遊星歯車機構とし、前段リングギヤを第１要素とし、前段キャリアと後段リングギヤとを連結して第２要素とし、後段キャリヤを第３要素とし、前段及び後段サンギヤを連結して第４要素としたことである。
【００１０】
請求項６に係る発明の構成上の特徴は、請求項１乃至請求項４のいずれかに記載の自動変速機において、前記変速用複式遊星歯車装置を前後段２組のシングルピニオン型の遊星歯車機構とし、前段サンギヤを第１要素とし、前段キャリヤと後段リングギヤとを連結して第２要素とし、前段リングギヤと後段キャリヤとを連結して第３要素とし、後段サンギヤを第４要素としたことである。
【００１１】
請求項７に係る発明の構成上の特徴は、請求項１乃至請求項４のいずれかに記載の自動変速機において、前記変速用複式遊星歯車装置をシングルピニオン型の前段遊星歯車機構とダブルピニオン型の後段遊星歯車機構とし、前段サンギヤを第１要素とし、前段及び後段キャリヤを連結して第２要素とし、前段及び後段リングギヤを連結して第３要素とし、後段サンギヤを第４要素としたことである。
【００１２】
請求項８に係る発明の構成上の特徴は、請求項１乃至請求項４のいずれかに記載の自動変速機において、前記変速用複式遊星歯車装置をシングルピニオン型の前段遊星歯車機構とダブルピニオン型の後段遊星歯車機構とし、前段リングギヤを第１要素とし、前段及び後段キャリヤを連結して第２要素とし、後段リングギヤを第３要素とし、前段及び後段サンギヤを連結して第４要素としたことである。
【００１３】
【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、入力軸の回転より回転数が小さい第１回転及び該第１回転より回転数が小さい第２回転を生成する歯車減速装置を設け、前記入力軸の回転を変速用複式遊星歯車装置の第２、第４要素に第２、第１制御クラッチにより選択的に伝達し、前記第１及び第２回転の一方を回転選択手段で選択して第１要素に選択的に伝達し、第１、第２要素の回転を第１及び第２制御ブレーキで選択的に規制し、第３要素を出力軸に連結したので、従来の自動変速機に最小限の変更を加えるだけで、入力軸の回転を適切に離間した前進６段以上、後退１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、前記歯車減速装置から前記第１回転及び前記第２回転の一方が前記回転選択手段によって選択されて前記第１要素に伝達された場合に、前記入力軸の回転が前記第２制御クラッチで前記第２要素に伝達されることにより前記入力軸の回転数より前記出力軸の回転数が大きい増速段を形成するので、高速段側のギヤ比を更に密にすることができて、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車速変化が小さくなり、良好なフィーリングを得ることができる。
【００１４】
上記のように構成した請求項２に係る発明においては、段付ピニオンを有する減速用複式遊星歯車装置の小径又は大径リングギヤの回転を選択的に規制してキャリヤに第１又は第２回転を選択的に生成させ、第１又は第２回転を変速用複式遊星歯車装置の第１要素に選択的に伝達し、入力軸の回転を第２、第４要素に第２、第１制御クラッチにより選択的に伝達し、第１、第２要素の回転を第１、第２制御ブレーキで選択的に規制し、第３要素を出力軸に連結したので、請求項１に記載した発明の効果に加え、従来の自動変速機の減速要用遊星歯車装置を段付ピニオンを有する複式のものとし、１個の制御ブレーキを追加するだけで、前進６段以上、後退１段のギヤ比で変速して出力軸に伝達することができるコンパクトな自動変速機を得ることができる。
【００１５】
上記のように構成した請求項３に係る発明においては、減速用複式遊星歯車装置の第１リングギヤ及びキャリヤの回転を第１、第２回転生成制御ブレーキで夫々選択的に規制して第２リングギヤに第１又は第２回転を選択的に生成させ、第１又は第２回転を変速用複式遊星歯車装置の第１要素に選択的に伝達し、入力軸の回転を第２、第４要素に第２、第１制御クラッチにより選択的に伝達し、第１、第２要素の回転を第１、第２制御ブレーキで選択的に規制し、第３要素を出力軸に連結したので、請求項１に記載した発明の効果に加え、従来の自動変速機の減速用遊星歯車装置を他のタイプのものとし、１個の制御ブレーキを追加するだけで、入力軸の回転を適切に離間した前進６段以上、後退１段のギヤ比で変速して出力軸に伝達することができるコンパクトな自動変速機を得ることができる。
【００１６】
上記のように構成した請求項４に係る発明においては、減速用歯車列によって入力回転、第１回転、第２回転を生成し、第１又は第２回転を第１、第２回転伝達制御クラッチにより変速用複式遊星歯車装置の第１要素に選択的に伝達し、入力回転を第２、第４要素に第２、第１制御クラッチにより選択的に伝達し、第１、第２要素の回転を第１、第２制御ブレーキで選択的に規制し、第３要素を出力軸に連結したので、請求項１に記載した発明の効果に加え、従来の自動変速機の減速用遊星歯車装置を簡単な減速用歯車列に変換し、１個の制御ブレーキを２個の制御クラッチに変更するだけで、入力軸の回転を適切に離間した前進６段以上、後退１段のギヤ比で変速して出力軸に伝達することができるコンパクトな自動変速機を得ることができる。
【００１７】
上記のように構成した請求項５に係る発明においては、変速用複式遊星歯車装置を前後段２組のシングルピニオン型の遊星歯車機構とし、前段リングギヤを第１要素とし、前段キャリアと後段リングギヤとを連結して第２要素とし、後段キャリヤを第３要素とし、前段及び後段サンギヤを連結して第４要素としたので、請求項１に記載の発明の効果に加え、簡単な構成でコンパクトな自動変速機を得ることができる。
【００１８】
上記のように構成した請求項６に係る発明においては、変速用複式遊星歯車装置を前後段２組のシングルピニオン型の遊星歯車機構とし、前段サンギヤを第１要素とし、前段キャリヤと後段リングギヤとを連結して第２要素とし、前段リングギヤと後段キャリヤとを連結して第３要素とし、後段サンギヤを第４要素としたので、従来の自動変速機に最小限の変更を加えるだけで、入力軸の回転を適切に離間した前進６段以上、後退１段のギヤ比で変速して出力軸に伝達することができる構造簡単でコンパクトな自動変速機を得ることができる。
【００１９】
上記のように構成した請求項７に係る発明においては、変速用複式遊星歯車装置をシングルピニオン型の前段遊星歯車機構とダブルピニオン型の後段遊星歯車機構とし、前段サンギヤを第１要素とし、前段及び後段キャリヤを連結して第２要素とし、前段及び後段リングギヤを連結して第３要素とし、後段サンギヤを第４要素としたので、請求項１に記載の発明の効果に加え、簡単な構成で全長の短いコンパクトな自動変速機を得ることができる。
【００２０】
上記のように構成した請求項８に係る発明においては、変速用複式遊星歯車装置をシングルピニオン型の前段遊星歯車機構とダブルピニオン型の後段遊星歯車機構とし、前段リングギヤを第１要素とし、前段及び後段キャリヤを連結して第２要素とし、後段リングギヤを第３要素とし、前段及び後段サンギヤを連結して第４要素としたので、従来の自動変速機に最小限の変更を加えるだけで、入力軸の回転を適切に離間した前進６段以上、後退１段のギヤ比で変速して出力軸に伝達することができる構造簡単で全長の短いコンパクトな自動変速機を得ることができる。
【００２１】
【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、２個のシングルピニオン型の遊星歯車機構５１，５２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、減速用複式遊星歯車装置１６は、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と噛合する小径ピニオン２３及び大径ピニオン２４からなる段付ピニオン２５、この段付きピニオン２５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、及び小径、大径ピニオン２３，２４と夫々噛合し共通軸線１３上に回転可能に支承された小径、大径リングギヤＲ１，Ｒ２から構成されている。入力軸１５はサンギヤＳ１，Ｓ２に連結されている。小径、大径リングギヤＲ１，Ｒ２をトランスミッションケース１２に夫々接続して選択的に回転を規制しキャリヤＣ１，Ｃ２に入力軸１５の回転より回転数の小さい第１回転及び第１回転より回転数の小さい第２回転を選択的に生成させる第１、第２回転生成制御ブレーキＢ−１，Ｂ−２が小径及び大径リングギヤＲ１，Ｒ２に夫々連結されている。このように減速用複式遊星歯車装置１６は、入力軸１５に連結され、入力軸１５の回転より回転数が小さい第１回転及びこの第１回転より回転数が小さい第２回転を生成する歯車減速装置４９を構成する。
【００２２】
変速用複式遊星歯車装置１７は、前後段２組のシングルピニオン型の遊星歯車機構５３，５４の前段キャリアＣ３及び後段リングギヤＲ４、前段及び後段サンギヤＳ３，Ｓ４を連結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前段及び後段サンギヤＳ３，Ｓ４を連結し、前段サンギヤＳ３に噛合するピニオン３３を支承し共通軸線１３上に回転可能に支承された前段キャリヤと後段リングギヤＲ４とを連結し、ピニオン３３に噛合する前段リングギヤＲ３を共通軸線１３上に回転可能に支承し、後段サンギヤＳ４と後段リングギヤＲ４とに噛合するピニオン３４を支承するキャリヤＣ４を共通軸線１３上に回転可能に支承して構成されている。
【００２３】
キャリヤＣ４には出力軸１８が連結されている。前段リングギヤＲ３は、減速用複式遊星歯車装置１６の共通のキャリヤＣ１，Ｃ２に連結されるとともに、第１制御ブレーキＢ−３によりトランスミッションケース１２に選択的に接続されて回転を規制されるようになっている。互いに連結された前段キャリヤＣ３と後段リングギヤＲ４とは、第２制御ブレーキＢ−４によりトランスミッションケース１２に選択的に接続されて回転を規制されるようになっている。入力軸１５の回転を変速用複式遊星歯車装置１７の前後段サンギヤＳ３，Ｓ４に選択的に伝達する第１制御クラッチＣ−１と、入力軸１５の回転を互いに連結された前段キャリヤＣ３と後段リングギヤＲ４に選択的に伝達する第２制御クラッチＣ−２が設けられている。そして、第１、第２回転生成制御ブレーキＢ−１，Ｂ−２は、減速用複式遊星歯車装置１６の共通キャリヤＣ１，Ｃ２に生成される第１及び第２回転の一方を選択して変速用複式遊星歯車装置１７の前段リングギヤＲ３に伝達する回転選択手段５０を構成する。
【００２４】
なお、流体トルクコンバータ１１のポンプインペラ４５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４６がオイルの反力を受け止めてトルクをタービン４７に発生するようになっている。入力軸１５はタービン４７に連結されている。４８はポンプインペラ４５とタービン４７とを直結するロックアップクラッチである。
【００２５】
以上のように構成された自動変速機１０においては、第１、第２制御クラッチＣ−１，Ｃ−２を選択的に係脱し、第1、第２制御ブレーキＢ−３，Ｂ−４及び第１、第２回転生成制御ブレーキＢ−１，Ｂ−２を選択的に作動して遊星歯車装置の要素の回転を規制することにより、前進８段、後退１段のギヤ比を成立することができる。図２において、各速度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、減速用複式遊星歯車装置１６のサンギヤＳ１、小径ピニオン２３、キャリヤＣ１及び小径リングギヤＲ１からなる減速用第１遊星歯車機構５１のギヤ比λ１が０．５０８、サンギヤＳ２、大径ピニオン２４、キャリヤＣ２及び大径リングギヤＲ２からなる減速用第２遊星歯車機構５２のギヤ比λ２が０．３０７、変速用複式遊星歯車装置１７の前段サンギヤＳ３、ピニオン３３、前段キャリヤＣ３及び前段リングギヤＲ３からなる変速用前段遊星歯車機構５３のギヤ比λ３が０．３７７、後段サンギヤＳ４、ピニオン３４、後段キャリヤＣ４及び後段リングギヤＲ４からなる変速用後段遊星歯車機構５４のギヤ比λ４が０．４２９である場合における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）がギヤ比欄に示されている。
【００２６】
シングルピニオン型の遊星歯車機構においては、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、各変速段におけるギヤ比は、式（１）に基づいて算出される。共通サンギヤＳ１，Ｓ２、前後段サンギヤＳ３，Ｓ４の歯数をＺｓ１，Ｚｓ２，Ｚｓ３，Ｚｓ４、小径、大径リングギヤＲ１，Ｒ２、前後段リングギヤＲ３，Ｒ４の歯数をＺｒ１，Ｚｒ２，Ｚｒ３，Ｚｒ４とすると、減速用第１、第２及び変速用第１、第２遊星歯車機構５１〜５４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）
【００２７】
第１、第２制御クラッチＣ−１，Ｃ−２を選択的に接続するとともに、第1、第２制御ブレーキＢ−３，Ｂ−４及び第１、第２回転生成制御ブレーキＢ−１，Ｂ−２を選択的に作動したとき、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を構成する減速用第１、第２遊星歯車機構５１，５２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。シングルピニオン型の第１遊星歯車機構５１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ａ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の縦線の反対側に間隔ａ×λ１だけ離して配置する。シングルピニオン型の第２遊星歯車機構５２についても同様に、キャリヤＣ２の縦線とサンギヤＳ２の縦線との間隔ｂ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ｂ×λ２だけ離して配置する。
【００２８】
変速用複式遊星歯車装置１７を構成する変速用第１、第２遊星歯車機構５３，５４では、前後段サンギヤＳ３，Ｓ４が共通し、前段キャリヤＣ３と後段リングギヤＲ４が共通するので、Ｓ３，Ｓ４及びＣ３，Ｒ４がそれぞれ付された各１本の縦線上に前後段サンギヤＳ３，Ｓ４、連結された前段キャリヤＣ３と後段リングギヤＲ４の速度比を表し、それぞれＲ３、Ｃ４が付された各１本の縦線上に前段リングギヤＲ３、後段キャリヤＣ４の速度比を表す。シングルピニオン型の変速用前段遊星歯車機構５３については、キャリヤＣ３の縦線とサンギヤＳ３の縦線との間隔ｃを１とみなし、リングギヤＲ３の縦線をキャリヤＣ３の縦線からサンギヤＳ３の縦線の反対側に間隔ｃ×λ３だけ離して配置する。シングルピニオン型の後段遊星歯車機構５４については、サンギヤＳ４の縦線とリングギヤＲ４の縦線との間隔ｄを１＋λ４とみなし、キャリヤＣ４の縦線をサンギヤＳ４の縦線からリングギヤＲ４の縦線と同じ側に間隔ｄ／（１＋λ４）だけ離して配置する。速度線図には、第１、第２回転生成制御ブレーキＢ−１，Ｂ−２、第１、第２制御クラッチＣ−１，Ｃ−２、第1、第２制御ブレーキＢ−３，Ｂ−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。
【００２９】
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要素としての前段リングギヤＲ３は減速用複式遊星歯車装置１６の共通のキャリヤＣ１，Ｃ２に連結されるとともに第１制御ブレーキＢ−３に連結され、第２要素としての互いに連結された前段キャリヤＣ３と後段リングギヤＲ４は第２制御クラッチＣ−２と第２制御ブレーキＢ−４とに並列に連結され、第３要素としての後段キャリヤＣ４は出力軸１８に連結され、第４要素としての互いに連結された前段及び後段サンギヤＳ３，Ｓ４は第１クラッチＣ−１に連結されている。
【００３０】
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１制御クラッチＣ−１、第２制御ブレーキＢ−４が作動され、入力軸１５と前後段サンギヤＳ３，Ｓ４が接続され、互いに連結された前段キャリヤＣ３と後段リングギヤＲ４が回転規制されるので、入力軸１５に入力された回転は、第１制御クラッチＣ−１を介して後段サンギヤＳ４に伝達され、回転を規制された後段リングギヤＲ４により反力を支持されてピニオン３４を公転させて後段キャリヤＣ４に伝達され、出力軸１８を第１変速段のギヤ比３．３３３で正転駆動する。
【００３１】
前進第２変速段の場合、第１制御クラッチＣ−１、第１制御ブレーキＢ−３が作動され、入力軸１５と前後段サンギヤＳ３，Ｓ４が接続され、前段リングギヤＲ３が回転規制されるので、入力軸１５に入力された回転は、第１制御クラッチＣ−１を介して前段サンギヤＳ３に伝達され、回転を規制された前段リングギヤＲ３により反力を支持されてピニオン３３を公転させて互いに連結された前段キャリヤＣ３と後段リングギヤＲ４を回転し、後段サンギヤＳ４と後段リングギヤＲ４との回転差に応じてピニオン３４を公転させて後段キャリヤＣ４と出力軸１８を第２変速段のギヤ比２．０３５で正転駆動する。
【００３２】
前進第３変速段の場合、第１制御クラッチＣ−１、第２回転生成制御ブレーキＢ−２が作動され、入力軸１５と前後段サンギヤＳ３，Ｓ４が接続され、大径リングギヤＲ２が回転規制されるので、入力軸１５に入力された回転は、第１制御クラッチＣ−１を介して前後段サンギヤＳ３，Ｓ４に伝達されるとともに、減速用複式遊星歯車装置１６のサンギヤＳ１，Ｓ２に伝達され、回転を規制された大径リングギヤＲ２により反力を支持されて段付ピニオン２５を公転させてキャリヤＣ１，Ｃ２に第２回転を生成し、変速用複式遊星歯車装置１７の前段リングギヤＲ３に第２回転を伝達し、前段サンギヤＳ３と前段リングギヤＲ３との回転差に応じてピニオン３３を公転させて前段キャリヤＣ３及び後段リングギヤＲ４を回転し、後段サンギヤＳ４と後段リングギヤＲ４との回転差に応じてピニオン３４を公転させて後段キャリヤＣ４と出力軸１８を第３変速段のギヤ比１．６３７で正転駆動する。
【００３３】
前進第４変速段の場合、第１制御クラッチＣ−１、第１回転生成制御ブレーキＢ−１が作動され、入力軸１５と前後段サンギヤＳ３，Ｓ４が接続され、小径リングギヤＲ１が回転規制されるので、入力軸１５に入力された回転は、第１制御クラッチＣ−１を介して前後段サンギヤＳ３，Ｓ４に伝達されるとともに、減速用複式遊星歯車装置１６のサンギヤＳ１，Ｓ２に伝達され、回転を規制された小径リングギヤＲ１により反力を支持されて段付ピニオン２５を公転させてキャリヤＣ１，Ｃ２に第１回転を生成し、変速用複式遊星歯車装置１７の前段リングギヤＲ３に第１回転を伝達し、前段サンギヤＳ３と前段リングギヤＲ３との回転差に応じてピニオン３３を公転させて前段キャリヤＣ３及び後段リングギヤＲ４を回転し、後段サンギヤＳ４と後段リングギヤＲ４との回転差に応じてピニオン３４を公転させて後段キャリヤＣ４と出力軸１８を第４変速段のギヤ比１．５０９で正転駆動する。
【００３４】
前進第５変速段の場合、第１、第２制御クラッチＣ−１，Ｃ−２が接続され、入力軸１５が前後段サンギヤＳ３，Ｓ４及び互いに連結された前段キャリヤＣ３と後段リングギヤＲ４に接続されるので、後段キャリヤＣ４も一体になって回転され、出力軸１８が第５変速段のギヤ比１．０００で正転駆動される。
【００３５】
前進第６変速段の場合、第２制御クラッチＣ−２、第１回転生成制御ブレーキＢ−１が作動され、入力軸１５が前段キャリヤＣ３と後段リングギヤＲ４に接続され、小径リングギヤＲ１が回転規制されるので、入力軸１５に入力された回転は、第２制御クラッチＣ−２を介して前段キャリヤＣ３と後段リングギヤＲ４に伝達されるとともに、減速用複式遊星歯車装置１６のサンギヤＳ１，Ｓ２に伝達され、回転を規制された小径リングギヤＲ１により反力を支持されて段付ピニオン２５を公転させてキャリヤＣ１，Ｃ２に第１回転を生成し、変速用複式遊星歯車装置１７の前段リングギヤＲ３に第１回転を伝達し、前段キャリヤＣ３と前段リングギヤＲ３との回転差に応じて前後段サンギヤＳ３，Ｓ４を回転し、後段サンギヤＳ４と後段リングギヤＲ４との回転差に応じてピニオン３４を公転させて後段キャリヤＣ４と出力軸１８を第６変速段のギヤ比０．６５４で正転駆動する。
【００３６】
前進第７変速段の場合、第２制御クラッチＣ−２、第２回転生成制御ブレーキＢ−２が作動され、入力軸１５が前段キャリヤＣ３と後段リングギヤＲ４に接続され、大径リングギヤＲ２が回転規制されるので、入力軸１５に入力された回転は、第２制御クラッチＣ−２を介して前段キャリヤＣ３と後段リングギヤＲ４に伝達されるとともに、減速用複式遊星歯車装置１６のサンギヤＳ１，Ｓ２に伝達され、回転を規制された大径リングギヤＲ２により反力を支持されて段付ピニオン２５を公転させてキャリヤＣ１，Ｃ２に第２回転を生成し、変速用複式遊星歯車装置１７の前段リングギヤＲ３に第２回転を伝達し、前段キャリヤＣ３と前段リングギヤＲ３との回転差に応じて前後段サンギヤＳ３，Ｓ４を回転し、後段サンギヤＳ４と後段リングギヤＲ４との回転差に応じてピニオン３４を公転させて後段キャリヤＣ４と出力軸１８を第７変速段のギヤ比０．６２１で正転駆動する。
【００３７】
前進第８変速段の場合、第２制御クラッチＣ−２、第１制御ブレーキＢ−３が作動され、入力軸１５が前段キャリヤＣ３と後段リングギヤＲ４に接続され、前段リングギヤＲ３が回転規制されるので、入力軸１５に入力された回転は、第２制御クラッチＣ−２を介して前段キャリヤＣ３と後段リングギヤＲ４に伝達され、回転を規制された前段リングギヤＲ３に反力を支持されて前後段サンギヤＳ３，Ｓ４を回転し、後段サンギヤＳ４と後段リングギヤＲ４との回転差に応じてピニオン３４を公転させて後段キャリヤＣ４と出力軸１８を第８変速段のギヤ比０．５５７で正転駆動する。
【００３８】
後退変速段の場合、第１回転生成制御ブレーキＢ−１及び第２制御ブレーキＢ−４が作動され、小径リングギヤＲ１、互いに連結された前段キャリヤＣ３と後段リングギヤＲ４が回転規制されるので、入力軸１５に入力された回転は、サンギヤＳ１，Ｓ２に伝達され、回転を規制された小径リングギヤＲ１により反力を支持されて段付ピニオン２５を公転させてキャリヤＣ１，Ｃ２に第１回転を生成し、変速用複式遊星歯車装置１７の前段リングギヤＲ３に第１回転を伝達し、回転を規制された前段キャリヤＣ３に支承されたピニオン３３を介して前後段サンギヤＳ３、Ｓ４を逆転し、後段リングギヤＲ４で反力を支持してピニオン３４を公転させて後段キャリヤＣ４と出力軸１８を後退変速段のギヤ比３．７２６で逆転駆動する。
【００３９】
入力軸１５に連結された減速用複式遊星歯車装置１６のサンギヤＳ１，Ｓ２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、各変速段における後段キャリヤＣ４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１段のギヤ比を得ることができる。
【００４０】
尚、第１の実施形態の前進第１変速段では、前段リングギヤＲ３が逆回転されるが、前段リングギヤＲ３に直接連結している共通キャリヤＣ１，Ｃ２も同様に逆転回転する。この際、共通サンギヤＳ１，Ｓ２は、入力軸１５と同じ回転数で回転しているため、小径リングギヤＲ１と大径リングギヤＲ２は、共通キャリヤＣ１，Ｃ２よりも大きく逆回転することになる。すると、共通サンギヤＳ１，Ｓ２と小径、大径リングギヤＲ１，Ｒ２との間には大きな相対回転差が生じ、その結果、両者に噛合する小径ピニオン２３及び大径ピニオン２４は、高速で自転する。
【００４１】
よって、この高速の自転を防止するために、第１の実施形態のキャリヤＣ１，Ｃ２と前段リングギヤＲ３との間、もしくは、共通サンギヤＳ１，Ｓ２と入力軸１５との間のどちらか一方に新たなクラッチを介在させて、選択的に連結可能とすることもできる。
【００４２】
この場合、この新たなクラッチは、キャリヤＣ１，Ｃ２から前段リングギヤＲ３へ動力が伝達される変速段、即ち、前進第３，４，６，７変速段及び後退変速段で係合させ、残りの変速段である前進第１，２，５，８変速段で開放させることができるが、上位の高速自転が顕著な前進第１変速段のみで開放させるようにしてもよい。
【００４３】
次に、第２の実施形態について、図４に基づいて説明する。第２の実施形態は、変速用複式遊星歯車装置１７、第１、第２制御クラッチＣ−１，Ｃ−２、第１、第２制御ブレーキＢ−３，Ｂ−４については、第１の実施形態と同じであるので、図面に同一符号を付けて説明を省略し、第１実施形態と異なる減速用遊星歯車装置６０のみについて説明する。
【００４４】
減速用複式遊星歯車装置６０は、図４に示すように、シングルピニオン型の遊星歯車機構６１及びダブルピニオン型の遊星歯車機構６２のサンギヤＳ１，Ｓ２、キャリヤＣ１，Ｃ２をそれぞれ連結、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た共通のサンギヤＳ１，Ｓ２、共通軸線１３上に回転可能に支承されてサンギヤＳ１，Ｓ２とロングピニオン６３を介して噛合する第１リングギヤＲ１、共通軸線１３上に回転可能に支承されてサンギヤＳ１，Ｓ２とロングピニオン６３及び中間ピニオン６４を介して噛合する第２リングギヤＲ２、ロングピニオン６３及び中間ピニオン６４を支承して共通軸線１３上に回転可能に支承された共通のキャリヤＣ１，Ｃ２から構成されている。共通のサンギヤＳ１，Ｓ２は入力軸１５に連結され、第２リングギヤＲ２は変速用複式遊星歯車装置１７の前段リングギヤＲ３に連結されている。
【００４５】
第１リングギヤＲ１には、第１リングギヤＲ１をトランスミッションケース１２に選択的に接続し回転を規制して第２リングギヤＲ２に入力軸１５の回転より回転数の小さい第１回転を生成させる第１回転生成制御ブレーキＢ−１が連結され、共通のキャリヤＣ１，Ｃ２には、キャリヤＣ１，Ｃ２をトランスミッションケース１２に選択的に接続し回転を規制して第２リングギヤＲ２に第１回転より回転数の小さい第２回転を生成させる第２回転生成制御ブレーキＢ−２が連結されている。
【００４６】
減速用複式遊星歯車装置６０は、入力軸１５に連結され、入力軸１５の回転より回転数が小さい第１回転及びこの第１回転より回転数が小さい第２回転を生成する歯車減速装置４９を構成する。そして、第１、第２回転生成制御ブレーキＢ−１，Ｂ−２は、減速用複式遊星歯車装置６０の第２リングギヤＲ２に生成される第１及び第２回転の一方を選択して変速用複式遊星歯車装置１７の前段リングギヤＲ３に伝達する回転選択手段５０を構成する。
【００４７】
第２実施形態においても、入力軸１５の回転を第１、第２制御クラッチＣ−１，Ｃ−２により変速用複式遊星歯車装置１７の前後段サンギヤＳ３，Ｓ４、互いに連結された前段キャリヤＣ３と後段リングギヤＲ４に選択的に伝達するとともに、第２リングギヤＲ２に生成された第１又は第２回転を前段リングギヤＲ３に選択的に伝達して入力軸１５の回転を前進８段、後退１段に変速することは、第１の実施形態の場合と同様であるので、詳細な説明は省略する。各速度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。
【００４８】
また図５には、減速用複式遊星歯車装置６０のサンギヤＳ１、ロングピニオン６３、キャリヤＣ１及びリングギヤＲ１からなる減速用第１遊星歯車機構６１のギヤ比λ１が０．５０８、サンギヤＳ２、ロングピニオン６３、中間ピニオン６４、キャリヤＣ２及びリングギヤＲ２からなる減速用第２遊星歯車機構６２のギヤ比λ２が０．３４１、変速用複式遊星歯車装置１７の前段サンギヤＳ３、ピニオン３３、前段キャリヤＣ３及び前段リングギヤＲ３からなる変速用前段遊星歯車機構５３のギヤ比λ３が０．３２５、後段サンギヤＳ４、ピニオン３４、後段キャリヤＣ４及び後段リングギヤＲ４からなる変速用後段遊星歯車機構５４のギヤ比λ４が０．３８１である場合における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）がギヤ比欄に示されている。
【００４９】
ダブルピニオン型の減速用第２遊星歯車機構６２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）で示される。各変速段におけるギヤ比は、この式（２）と前述のシングルピニオン型の遊星歯車機構における関係式（１）とに基づいて算出される。シングルピニオン型の遊星歯車機構においても、サンギヤＳ２の歯数をＺｓ２、第２リングギヤＲ２の歯数をＺｒ２とすると、減速用第２遊星歯車機構６２のギヤ比はλ２＝Ｚｓ２／Ｚｒ２である。
Ｎｒ＝（１−λ）Ｎｃ＋λＮｓ・・・（２）
【００５０】
第１、第２制御クラッチＣ−１，Ｃ−２を選択的に接続するとともに第1、第２制御ブレーキＢ−３，Ｂ−４を選択的に作動し、第１、第２回転生成制御ブレーキＢ−１，Ｂ−２を選択的に作動したとき、減速用複式遊星歯車装置６０及び変速用複式遊星歯車装置１７の各要素の速度比は、図６に示す速度線図のようになる。ダブルピニオン型の減速用第２遊星歯車機構６２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ｂ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の縦線と同じ側に間隔ｂ×λ２だけ離して配置する。速度線図には、第１、第２回転生成制御ブレーキＢ−１，Ｂ−２、第１、第２制御クラッチＣ−１，Ｃ−２、第1、第２制御ブレーキＢ−３，Ｂ−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。図６の速度線図に示すように、変速用複式遊星歯車装置１７の前段リングギヤＲ３が第１要素、互いに連結された前段キャリヤＣ３と後段リングギヤＲ４が第２要素、後段キャリヤＣ４が第３要素、互いに連結された前段及び後段サンギヤＳ３，Ｓ４が第４要素となる。変速用複式遊星歯車装置１７の各速度段における制御クラッチ、制御ブレーキの作動状態は、図２に示す第１実施形態の場合と同一であるので、作動の詳細な説明は省略する。
【００５１】
尚、第２の実施形態において、第１の実施形態と同様の理由から、第２リングギヤＲ２と前段リングギヤＲ３との間、もしくは、入力軸１５と共通サンギヤＳ１，Ｓ２との間のどちらか一方に新たなクラッチを介在させることもできる。
【００５２】
次に、歯車減速装置４９を減速用歯車列で構成した第３実施形態について説明する。第３実施形態は、変速用複式遊星歯車装置１７、第１、第２制御クラッチＣ−１，Ｃ−２、第１、第２制御ブレーキＢ−３，Ｂ−４については、第１実施形態と同じであるので、図７に同一符号を付けて説明を省略し、第１実施形態と異なる減速用歯車列６５及び減速用歯車列６５と変速用複式遊星歯車装置１７との接続関係について説明する。自動変速機１０のトランスミッションケース１２に回転可能に軸承された入力軸６６に流体トルクコンバータ１１のタービン４７が連結され、この入力軸６６に大径、中径及び小径歯車６７，６８，６９が固定されている。大径歯車６７と噛合する同径の第１歯車７０が変速用複式遊星歯車装置１７の軸線１３上に回転可能に支承され、中径、小径歯車６８，６９とそれぞれ噛合する第２、第３歯車７１，７２が軸線１３上に回転可能に支承されている。これにより第１歯車７０は入力軸６６の回転と同一回転数で回転し、第２歯車７１は入力軸６６の回転より回転数が小さい第１回転で回転し、第３歯車７２は第１回転より回転数が小さい第２回転で回転する。第1歯車７０は、第１制御クラッチＣ−１により変速用複式遊星歯車装置１７の前後段サンギヤＳ１，Ｓ２に係脱可能に連結され、第２制御クラッチＣ−２により前段キャリヤＣ３と後段リングギヤＲ４に係脱可能に連結されるようになっている。第２、第３歯車７１，７２は、第１、第２回転伝達制御クラッチＣ−４，Ｃ−３により前段リングギヤＲ３に係脱可能に連結されるようになっている。第１、第２回転伝達制御クラッチＣ−４，Ｃ−３は、減速用歯車列６５により生成される第１及び第２回転の一方を選択して変速用複式遊星歯車装置１７の前段リングギヤＲ３に伝達する回転選択手段５０を構成する。
【００５３】
変速用複式遊星歯車装置１７の速度線図は、図８に示すように、第１実施形態の場合と同一で、前段リングギヤＲ３が第１要素、前段キャリヤＣ３と後段リングギヤＲ４が第２要素、後段キャリヤＣ４が第３要素、前後段サンギヤＳ３，Ｓ４が第４要素となる。変速用複式遊星歯車装置１７の各速度段における制御クラッチ、制御ブレーキの作動状態は、第１、第２回転生成制御ブレーキＢ−１，Ｂ−２に替えて第１、第２回転伝達制御クラッチＣ−４，Ｃ−３を作動させれば図２に示す第１実施形態の場合と同一であるので、作動の詳細な説明は省略する。
【００５４】
次に、歯車減速装置４９に第１実施形態と同一の減速用複式遊星歯車装置１６を使用し、変速用複式遊星歯車装置を上記実施形態と異なる複式遊星歯車装置で構成した他の実施形態について説明する。減速用複式遊星歯車装置１６は、第１実施形態のものと同一であるので、図面に同一符号を付して説明を省略する。
【００５５】
第４実施形態においては、変速用複式遊星歯車装置７５は、図９に示すように、前後段２組のシングルピニオン型の遊星歯車機構７６，７７で構成され、ピニオン７８を支承する前段キャリアＣ３と後段リングギヤＲ４が連結され、ピニオン７９を支承する後段キャリヤＣ４と前段リングギヤＲ３が連結されている。後段キャリヤＣ４に出力軸１８が連結されている。前段サンギヤＳ３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、後段サンギヤＳ４は第１制御クラッチＣ−１を介して入力軸１５に係脱可能に連結され、前段キャリヤＣ３と後段リングギヤＲ４は第２制御クラッチＣ−２を介して入力軸１５に係脱可能に連結されている。前段サンギヤＳ３は第1制御ブレーキＢ−３に連結され、前段キャリヤＣ３と後段リングギヤＲ４は第２制御ブレーキＢ−４に連結されて回転を選択的に規制されるようになっている。
【００５６】
尚、第４の実施形態において、第１の実施形態と同様の理由から、キャリヤＣ１，Ｃ２と前段サンギヤＳ３との間、もしくは、入力軸１５と共通サンギヤＳ１，Ｓ２との間のどちらか一方に新たなクラッチを介在させることもできる。
【００５７】
図１０の速度線図に示すように、前段サンギヤＳ３が第１要素、互いに連結された前段キャリヤＣ３と後段リングギヤＲ４が第２要素、互いに連結された前段リングギヤＲ３と後段キャリヤＣ４が第３要素、後段サンギヤＳ４が第４要素となる。各速度段における制御クラッチ、制御ブレーキの作動状態は、図２に示す第１実施形態のものと同一であるので、作動の詳細な説明は省略する。
【００５８】
第５実施形態においては、図１１に示すように、変速用複式遊星歯車装置８０は、シングルピニオン型の前段遊星歯車機構８１及びダブルピニオン型の後段遊星歯車機構８２の前段及び後段キャリヤＣ３，Ｃ４、前段及び後段リングギヤＲ３，Ｒ４をそれぞれ連結、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前段及び後段サンギヤＳ３，Ｓ４、前段サンギヤＳ３に直接噛合するとともに後段サンギヤＳ４に中間ピニオン８３を介して噛合するロングピニオン８４、ロングピニオン８４及び中間ピニオン８３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前段及び後段キャリヤＣ３，Ｃ４、及びロングピニオン８４と噛合し共通軸線１３上に回転可能に支承された前段及び後段リングギヤＲ３，Ｒ４から構成されている。リングギヤＲ３，Ｒ４は、出力軸１８に連結されている。そして、前段サンギヤＳ３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、後段サンギヤＳ４は第１制御クラッチＣ−１を介して入力軸１５に係脱可能に連結され、前後段キャリヤＣ３，Ｃ４は第２制御クラッチＣ−２を介して入力軸１５に係脱可能に連結されるようになっている。前段サンギヤＳ３は第１制御ブレーキＢ−３に連結され、前後段キャリヤＣ３，Ｃ４は第２制御ブレーキＢ−４に連結されて回転を選択的に規制されるようになっている。
【００５９】
尚、第５の実施形態において、第１の実施形態と同様の理由から、キャリヤＣ１，Ｃ２と前段サンギヤＳ３との間、もしくは、入力軸１５と共通サンギヤＳ１，Ｓ２との間のどちらか一方に新たなクラッチを介在させることもできる。
【００６０】
図１２の速度線図に示すように、前段サンギヤＳ３が第１要素、互いに連結された前段及び後段キャリヤＣ３，Ｃ４が第２要素、互いに連結された前段及び後段リングギヤＲ３，Ｒ４が第３要素、後段サンギヤＳ４が第４要素となる。変速用複式遊星歯車装置８０の各速度段における制御クラッチ、制御ブレーキの作動状態は、図２に示す第１実施形態のものと同一であるので、作動の詳細な説明は省略する。
【００６１】
第６実施形態においては、図１３に示すように、変速用複式遊星歯車装置９０は、シングルピニオン型の前段遊星歯車機構９１及びダブルピニオン型の後段遊星歯車機構９２の前段及び後段サンギヤＳ３，Ｓ４、前段及び後段キャリヤＣ３，Ｃ４をそれぞれ連結、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た共通のサンギヤＳ３，Ｓ４、共通軸線１３上に回転可能に支承されてサンギヤＳ３，Ｓ４とロングピニオン９３を介して噛合するリングギヤＲ３、共通軸線１３上に回転可能に支承されてサンギヤＳ３，Ｓ４とロングピニオン９３及び中間ピニオン９４を介して噛合するリングギヤＲ４、ロングピニオン９３及び中間ピニオン９４を支承して共通軸線１３上に回転可能に支承された共通のキャリヤＣ３，Ｃ４から構成されている。リングギヤＲ４に出力軸１８が連結されている。そして、前段リングギヤＲ３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、前後段サンギヤＳ３，Ｓ４は第１制御クラッチＣ−１を介して入力軸１５に係脱可能に連結され、前後段キャリヤＣ３，Ｃ４は第２制御クラッチＣ−２を介して入力軸１５に係脱可能に連結されるようになっている。前段リングギヤＲ３は第１制御ブレーキＢ−３に連結され、前後段キャリヤＣ３，Ｃ４は第２制御ブレーキＢ−４に連結されて回転を選択的に規制されるようになっている。
【００６２】
尚、第６の実施形態において、第１の実施形態と同様の理由から、キャリヤＣ１，Ｃ２と前段リングギヤＲ３との間、もしくは、入力軸１５と共通サンギヤＳ１，Ｓ２との間のどちらか一方に新たなクラッチを介在させることもできる。
【００６３】
図１４の速度線図に示すように、前段リングギヤＲ３が第１要素、互いに連結された前段及び後段キャリヤＣ３，Ｃ４が第２要素、後段リングギヤＲ４が第３要素、互いに連結された前段及び後段サンギヤＳ３，Ｓ４が第４要素となる。各速度段における制御クラッチ、制御ブレーキの作動状態は、図２に示す第１実施形態のものと同一であるので、作動の詳細な説明は省略する。
【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。
【図７】 第３実施形態を示すスケルトン図である。
【図８】 第３実施形態の各速度段における変速用複式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９】 第４実施形態を示すスケルトン図である。
【図１０】 第４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１１】 第５実施形態を示すスケルトン図である。
【図１２】 第５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１３】 第６実施形態を示すスケルトン図である。
【図１４】 第６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、１５，６６・・・入力軸、１６，６０・・・減速用複式遊星歯車装置、１７，７５，８０，９０・・・変速用複式遊星歯車装置、１８・・・出力軸、２３・・・小径ピニオン、２４・・・大径ピニオン、２５・・・段付ピニオン、３３，３４，７８，７９・・・ピニオン、６４，８３，９４・・・中間ピニオン、６３，８４，９３・・・ロングピニオン、４９・・・歯車減速装置、５０・・・回転選択手段、５１，５２，５３，５４，６１，７６，７７，８１，９１・・・シングルピニオン型の遊星歯車機構、６２，８２，９２・・・ダブルピニオン型の遊星歯車機構、６５・・・減速用歯車列、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、Ｃ−１，Ｃ−２・・・第１、第２制御クラッチ、Ｃ−３，Ｃ−４・・・第２、第１回転伝達制御クラッチ、Ｂ−１，Ｂ−２・・・第１、第２回転生成制御ブレーキ、Ｂ−３，Ｂ−４・・・第１、第２制御ブレーキ。[0001]
BACKGROUND OF THE INVENTION
The present invention automatically engages / disengages a control clutch and a control brake connected to each element of a shift type planetary gear device connected to an input shaft, shifts the rotation of the input shaft to a plurality of stages, and transmits the rotation to the output shaft. It relates to a transmission.
[0002]
[Prior art]
Carrier for supporting a pair of pinions that mesh with each other, a sun gear that meshes with one of the pair of pinions, a double pinion type reduction planetary gear device having a ring gear that meshes with the other, a front stage and a rear stage sun gear, a front stage and a rear stage ring gear, It has a front stage carrier that supports a pinion that meshes with the front stage sun gear and the front stage ring gear, and a rear stage carrier that supports the pinion that meshes with the rear stage sun gear and the rear stage ring gear. The transmission type planetary gear unit having a gear connected to each other is provided on a common axis, the output shaft is connected to the rear carrier, and the rotation of the input shaft is connected to the front and rear sun gears connected to each other, the front carrier and the rear ring gear connected to each other. 1, selectively transmitted by the second control clutch, planetary gear set for reduction The rotation of the ring gear is transmitted to the front ring gear, and the rotations of the carrier of the reduction planetary gear device, the ring gear, and the front carrier and the rear ring gear connected to each other of the shift type planetary gear device are first, second, and second. Japanese Patent Application Laid-Open No. 2-129446 discloses an automatic transmission that selectively restricts by three control brakes and establishes a gear ratio of five forward speeds and one reverse speed.
[0003]
[Problems to be solved by the invention]
The conventional automatic transmission can smoothly and stably switch the gear ratio between five forward speeds and one reverse speed. However, in recent years, there has been an automatic transmission capable of establishing a gear ratio of six or more forward steps that are appropriately separated 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 demanded.
[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 six or more forward stages and one reverse stage that are appropriately separated.
[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 an input shaft, a first rotation coupled to the input shaft and having a lower rotational speed than the rotation of the input shaft, and the first rotation A gear reduction device that generates a second rotation having a smaller rotation speed, and first, second, third, and third elements respectively corresponding to the four elements arranged in order at intervals corresponding to the gear ratio in the velocity diagram. A double planetary gear unit for transmission having a fourth element, a second and first control clutch for selectively transmitting the rotation of the input shaft to the second and fourth elements, and one of the first and second rotations. Rotation selection means for selecting and transmitting to the first element, first and second control brakes for selectively restricting rotation of the first and second elements, and an output shaft connected to the third element The shift type planetary gear unit includes the first reduction gear unit from the gear reduction unit. Rolling as well as Second rotation One of When selected by the rotation selection means and transmitted to the first element In addition, The rotation of the input shaft is transmitted to the fourth element by the first control clutch, thereby forming a speed reduction stage in which the rotational speed of the output shaft is smaller than the rotational speed of the input shaft. When one of the first rotation and the second rotation is selected from the gear reduction device by the rotation selection means and transmitted to the first element, The rotation of the input shaft is transmitted to the second element by the second control clutch, thereby forming a speed increasing stage in which the rotational speed of the output shaft is larger than the rotational speed of the input shaft.
[0006]
According to a second aspect of the present invention, in the automatic transmission according to the first aspect, a stepped pinion comprising a sun gear coupled to the input shaft, a small-diameter pinion meshing with the sun gear, and a large-diameter pinion is supported. The gear reduction device is constituted by a reduction planetary gear device comprising a carrier, a small-diameter ring gear that meshes with a small-diameter pinion, and a large-diameter ring gear that meshes with the large-diameter pinion. And a first planetary gear device for selectively generating rotations of the small-diameter and large-diameter ring gears to selectively generate the first and second rotations. That is, the rotation selection means is configured by a rotation generation control brake.
[0007]
According to a third aspect of the present invention, in the automatic transmission according to the first aspect, the sun gear coupled to the input shaft, the long pinion meshing with the sun gear, and the intermediate pinion meshing with the long pinion are supported. And the second ring gear of the double planetary gear unit for reduction. The reduction planetary gear unit is composed of a carrier that rotates, a first ring gear that meshes with the long pinion, and a second ring gear that meshes with the intermediate pinion. Is coupled to the first element of the speed-changing double planetary gear unit, and the first ring gear and the carrier are selectively regulated to selectively generate the first and second rotations in the second ring gear, respectively. The first and second rotation generation control brakes constitute the rotation selection means.
[0008]
The constitutional feature of the invention according to claim 4 is the automatic transmission according to claim 1, wherein the automatic transmission according to claim 1 can rotate concentrically with the large-diameter, medium-diameter, small-diameter gear fixed to the input shaft, and the planetary gear device for transmission. And an input gear that meshes with the large diameter, medium diameter, and small diameter gears respectively to generate an input rotation, a first rotation having a rotation speed smaller than the input rotation, and a second rotation having a rotation speed smaller than the first rotation. The reduction gear train comprising the first gear and the second gear constitutes the gear reduction device, the input gear is connected to the second and first control clutches, and the input rotation is changed to the shift type planetary gear device. The first and second rotation transmission control clutches selectively transmit to the second and fourth elements and detachably connect the first and second gears to the first element, respectively. A rotation selection means for selecting one of the rotations and transmitting it to the first element; It is.
[0009]
A structural feature of the invention according to claim 5 is that in the automatic transmission according to any one of claims 1 to 4, the double planetary gear device for shifting is composed of two sets of single-pinion type planetary gears in the front and rear stages. The mechanism is that the front ring gear is the first element, the front carrier and the rear ring gear are connected as the second element, the rear carrier is the third element, and the front and rear sun gears are connected as the fourth element. .
[0010]
The structural feature of the invention according to claim 6 is the automatic transmission according to any one of claims 1 to 4, wherein the double planetary gear device for shifting is replaced with two sets of single-pinion type planetary gears in the front and rear stages. The mechanism is such that the front stage sun gear is the first element, the front stage carrier and the rear stage ring gear are connected as the second element, the front stage ring gear and the rear stage carrier are connected as the third element, and the rear stage sun gear is the fourth element. It is.
[0011]
According to a seventh aspect of the present invention, there is provided an automatic transmission according to any one of the first to fourth aspects, wherein the double planetary gear unit for transmission is replaced with a single-pinion type front planetary gear mechanism and a double-pinion. A rear planetary gear mechanism of the mold, the front stage sun gear is the first element, the front stage and rear stage carrier are connected as the second element, the front stage and rear stage ring gear are connected as the third element, and the rear stage sun gear is the fourth element That is.
[0012]
According to an eighth aspect of the present invention, there is provided an automatic transmission according to any one of the first to fourth aspects, wherein the double planetary gear device for transmission is replaced with a single pinion type front planetary gear mechanism and a double pinion. A rear planetary gear mechanism of the type, the front ring gear is the first element, the front and rear carriers are connected as the second element, the rear ring gear is the third element, and the front and rear stage sun gears are connected as the fourth element. That is.
[0013]
[Operation and effect of the invention]
In the invention according to claim 1 configured as described above, there is provided a gear reduction device that generates a first rotation having a rotation speed smaller than the rotation of the input shaft and a second rotation having a rotation speed smaller than the first rotation, The rotation of the input shaft is selectively transmitted to the second and fourth elements of the shift type planetary gear unit by the second and first control clutches, and one of the first and second rotations is selected by the rotation selecting means. Since it selectively transmits to the first element, the rotation of the first and second elements is selectively restricted by the first and second control brakes, and the third element is connected to the output shaft, the conventional automatic transmission It is possible to provide an automatic transmission that can change the speed of the input shaft with a gear ratio of 6 or more forward speeds and 1 reverse speed and transmit them to the output shaft with minimal changes. . Further, the first rotation from the gear reduction device as well as Second rotation One of When selected by the rotation selection means and transmitted to the first element In addition, Since the rotation of the input shaft is transmitted to the second element by the second control clutch, a speed increasing stage in which the rotational speed of the output shaft is larger than the rotational speed of the input shaft is formed. The ratio can be made denser, the engine performance can be optimally extracted in the high speed range of the vehicle speed, and the change of the gear ratio at the time of the gear change can be reduced. Obtainable.
[0014]
In the invention according to claim 2 configured as described above, the rotation of the small-diameter or large-diameter ring gear of the double planetary gear unit for reduction having a stepped pinion is selectively restricted, and the first or second rotation is performed on the carrier. The first or second rotation is selectively transmitted to the first element of the shift type planetary gear device, and the rotation of the input shaft is transmitted to the second and fourth elements by the second and first control clutches. Since the transmission is selectively performed, the rotation of the first and second elements is selectively restricted by the first and second control brakes, and the third element is connected to the output shaft, the effect of the invention described in claim 1 is achieved. In addition, the conventional planetary gear unit for speed reduction of an automatic transmission is a compound type having a stepped pinion, and by adding one control brake, the gear ratio is changed at a gear ratio of 6 or more forwards and 1 reverse. To obtain a compact automatic transmission that can transmit to the output shaft Door can be.
[0015]
In the invention according to claim 3 configured as described above, the rotation of the first ring gear and the carrier of the double planetary gear unit for reduction is selectively restricted by the first and second rotation generation control brakes, respectively. To selectively generate the first or second rotation, selectively transmit the first or second rotation to the first element of the double planetary gear device for shifting, and rotate the input shaft to the second or fourth element. The second and first control clutches selectively transmit, the first and second elements rotate selectively with the first and second control brakes, and the third element is connected to the output shaft. In addition to the effects of the invention described in item 1, the conventional planetary gear unit for reduction of an automatic transmission is of another type, and only one control brake is added, and the rotation of the input shaft is appropriately separated forward. Shift to a gear ratio of 6 steps or more and reverse 1 step and transmit to the output shaft. It is possible to obtain a compact automatic transmission which can.
[0016]
In the invention according to claim 4 configured as described above, the input rotation, the first rotation, and the second rotation are generated by the reduction gear train, and the first and second rotations are converted into the first and second rotation transmission control clutches. To selectively transmit the input rotation to the first element of the shift type planetary gear unit, to selectively transmit the input rotation to the second and fourth elements by the second and first control clutches, and to rotate the first and second elements. Is controlled selectively by the first and second control brakes, and the third element is connected to the output shaft. Therefore, in addition to the effect of the invention described in claim 1, a planetary gear unit for reduction of a conventional automatic transmission is provided. By simply converting to a gear train for reduction and changing one control brake to two control clutches, the input shaft can be rotated at a gear ratio of 6 or more forward speeds and 1 reverse speed, with appropriate separation. To obtain a compact automatic transmission that can transmit to the output shaft. That.
[0017]
In the invention according to claim 5 configured as described above, the speed-changing double planetary gear device is a single pinion type planetary gear mechanism having two sets of front and rear stages, the front ring gear is the first element, the front carrier and the rear ring gear are To the second element, the rear carrier to the third element, and the front and rear sun gears to be the fourth element. Therefore, in addition to the effect of the invention of claim 1, the structure is simple and compact. An automatic transmission can be obtained.
[0018]
In the invention according to claim 6 configured as described above, the speed change compound planetary gear device is a single pinion type planetary gear mechanism of two sets of front and rear stages, the front stage sun gear is the first element, the front stage carrier, the rear stage ring gear, Is connected to the second element, the front ring gear and the rear carrier are connected to form the third element, and the rear sun gear is set to the fourth element, so that the input can be made with minimal changes to the conventional automatic transmission. It is possible to obtain an automatic transmission with a simple structure and a compact structure capable of shifting the rotation of the shaft at a gear ratio of 6 or more forward speeds and appropriately 1 gear speed and transmitting them to the output shaft.
[0019]
In the invention according to claim 7 configured as described above, the shift type double planetary gear device is a single pinion type front planetary gear mechanism and a double pinion type rear planetary gear mechanism, the front stage sun gear is the first element, and the front stage In addition to the effect of the invention of claim 1, a simple structure is obtained because the second stage carrier is connected to form the second element, the first stage and rear ring gears are connected to form the third element, and the rear stage sun gear is set to the fourth element. A compact automatic transmission with a short overall length can be obtained.
[0020]
In the invention according to claim 8 configured as described above, the shift type double planetary gear device is a single pinion type front planetary gear mechanism and a double pinion type rear planetary gear mechanism, the front ring gear is the first element, and the front stage And the rear stage carrier is connected as the second element, the rear stage ring gear is the third element, the front stage and the rear stage sun gear are connected as the fourth element, so only a minimal change is made to the conventional automatic transmission, A compact automatic transmission having a simple structure and a short overall length can be obtained in which the rotation of the input shaft can be shifted at a gear ratio of 6 or more forward speeds and 1 reverse speed geared appropriately and transmitted to the output shaft.
[0021]
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 double planetary gear unit 16 for reduction is configured by connecting and sharing the sun gears S1 and S2 and the carriers C1 and C2 of the two single pinion type planetary gear mechanisms 51 and 52. That is, the reduction type planetary gear unit 16 includes a stepped pinion 25 including a common sun gear S1 and S2 rotatably supported on a common axis 13, and a small-diameter pinion 23 and a large-diameter pinion 24 that mesh with the sun gears S1 and S2. The stepped pinion 25 is rotatably supported and meshed with the common carriers C1 and C2 rotatably supported on the common axis 13 and the small and large diameter pinions 23 and 24, respectively, and can rotate on the common axis 13. Are comprised of small and large diameter ring gears R1, R2. The input shaft 15 is connected to the sun gears S1 and S2. The small-diameter and large-diameter ring gears R1 and R2 are connected to the transmission case 12 to selectively restrict the rotation, and the carriers C1 and C2 have a first rotation that is smaller than the rotation of the input shaft 15 and a rotation that is smaller than the first rotation. First and second rotation generation control brakes B-1 and B-2 for selectively generating a small second rotation are connected to the small diameter and large diameter ring gears R1 and R2, respectively. Thus, the reduction type planetary gear unit 16 is connected to the input shaft 15 and generates a first rotation having a rotation speed smaller than the rotation of the input shaft 15 and a second rotation having a rotation speed smaller than the first rotation. A device 49 is configured.
[0022]
The shift type planetary gear unit 17 is configured by connecting the front carrier C3 and the rear ring gear R4, and the front and rear sun gears S3 and S4 of the two sets of front and rear single-pinion type planetary gear mechanisms 53 and 54. That is, the front stage and rear stage sun gears S3 and S4 rotatably supported on the common axis 13 are connected, the pinion 33 meshing with the front stage sun gear S3 is supported, and the front stage carrier and the rear stage rotatably supported on the common axis line 13 are supported. The ring gear R4 is connected, the front ring gear R3 meshing with the pinion 33 is rotatably supported on the common axis 13, and the carrier C4 bearing the pinion 34 meshing with the rear sun gear S4 and the rear ring gear R4 is mounted on the common axis 13. It is configured to be rotatably supported.
[0023]
An output shaft 18 is connected to the carrier C4. The front ring gear R3 is coupled to the common carriers C1 and C2 of the double planetary gear unit 16 for reduction, and is selectively connected to the transmission case 12 by the first control brake B-3 so that rotation is restricted. It has become. The front stage carrier C3 and the rear stage ring gear R4 connected to each other are selectively connected to the transmission case 12 by the second control brake B-4 to restrict the rotation. A first control clutch C-1 for selectively transmitting the rotation of the input shaft 15 to the front and rear sun gears S3, S4 of the shift type planetary gear unit 17, and the front carrier C3 and the rear stage coupled to each other for the rotation of the input shaft 15. A second control clutch C-2 that selectively transmits to the ring gear R4 is provided. The first and second rotation generation control brakes B-1 and B-2 select one of the first and second rotations generated on the common carriers C1 and C2 of the double planetary gear unit 16 for speed reduction and change the speed. The rotation selection means 50 which transmits to the front | former stage ring gear R3 of the compound planetary gear apparatus 17 for an object is comprised.
[0024]
The pump impeller 45 of the fluid torque converter 11 is rotationally driven by an engine (not shown) to send out oil, and the stator 46 receives the reaction force of the oil and generates torque in the turbine 47. The input shaft 15 is connected to the turbine 47. A lockup clutch 48 directly connects the pump impeller 45 and the turbine 47.
[0025]
In the automatic transmission 10 configured as described above, the first and second control clutches C-1 and C-2 are selectively engaged and disengaged, and the first and second control brakes B-3 and B-4 and By selectively actuating the first and second rotation generation control brakes B-1 and B-2 to restrict the rotation of the elements of the planetary gear unit, a gear ratio of eight forward speeds and one reverse speed is established. Can do. In FIG. 2, when each control clutch and control brake corresponding to each speed stage is marked with a black circle, this indicates that the control clutch is in the connected state, and that the control brake is in the rotation restricted state. Further, FIG. 2 shows that the gear ratio λ1 of the first planetary gear mechanism 51 for reduction composed of the sun gear S1, the small-diameter pinion 23, the carrier C1 and the small-diameter ring gear R1 of the double planetary gear device 16 for reduction is 0.508, The gear ratio λ2 of the second planetary gear mechanism 52 for reduction comprising the large-diameter pinion 24, the carrier C2 and the large-diameter ring gear R2 is 0.307, the front stage sun gear S3, the pinion 33, the front stage carrier C3, The gear ratio λ3 of the shifting front planetary gear mechanism 53 including the front ring gear R3 is 0.377, and the gear ratio λ4 of the shifting rear planetary gear mechanism 54 including the rear sun gear S4, the pinion 34, the rear carrier C4, and the rear ring gear R4 is 0. .429, the gear ratio (the rotational speed of the input shaft 15 / the rotational speed of the output shaft 18) at each gear position is It is shown in the specific field.
[0026]
In the single-pinion type planetary gear mechanism, the relationship among the rotation speed Ns of the sun gear, the rotation speed Nc of the carrier, the rotation speed Nr of the ring gear, and the gear ratio λ of the planetary gear mechanism is expressed by equation (1). The gear ratio in the stage is calculated based on the formula (1). The number of teeth of the common sun gears S1, S2, the front and rear stage sun gears S3, S4 is Zs1, Zs2, Zs3, Zs4, the small diameter, large diameter ring gears R1, R2, and the number of teeth of the front and rear stage ring gears R3, R4 are Zr1, Zr2, Zr3, Zr4. Then, the gear ratio of the first, second, and first and second planetary gear mechanisms 51 to 54 for reduction is λ1 = Zs1 / Zr1, λ2 = Zs2 / Zr2, λ3 = Zs3 / Zr3, λ4 = Zs4 / Zr4.
Nr = (1 + λ) Nc−λNs (1)
[0027]
While selectively connecting the first and second control clutches C-1 and C-2, the first and second control brakes B-3 and B-4, and the first and second rotation generation control brakes B-1 and B-2, When B-2 is selectively operated, the speed ratio of each element of the speed reduction double planetary gear unit 16 and the speed change double planetary gear unit 17 is as shown in the speed diagram of 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 first and second planetary gear mechanisms 51 and 52 for reduction constituting the double planetary gear device 16 for reduction, the sun gears S1 and S2 and the carriers C1 and C2 are common to each other. Therefore, S1, S2 and C1, C2 are attached respectively. The speed ratios of the common sun gears S1 and S2 and the common carriers C1 and C2 are represented on each one vertical line, and the speeds of the ring gears R1 and R2 are represented on each vertical line to which R1 and R2 are respectively attached. Represents the ratio. For the single pinion type first planetary gear mechanism 51, the distance a 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 changed from the vertical line of the carrier C1 to the vertical line of the sun gear S1. Are spaced apart by a distance a × λ1. Similarly, for the single pinion type second planetary gear mechanism 52, the distance b 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 changed from the vertical line of the carrier C2 to the sun gear S2. They are arranged on the opposite side of the vertical line by a distance of b × λ2.
[0028]
In the first and second planetary gear mechanisms 53 and 54 for transmission constituting the double planetary gear device 17 for transmission, the front and rear stage sun gears S3 and S4 are common, and the front stage carrier C3 and the rear stage ring gear R4 are common, so S3 and S4 And the speed ratios of the front and rear stage sun gears S3 and S4, the connected front stage carrier C3 and the rear stage ring gear R4 on each one vertical line to which C3 and R4 are respectively attached, and one each having R3 and C4 respectively. The speed ratio of the front ring gear R3 and the rear carrier C4 is shown on the vertical line. With regard to the single-pinion type front planetary gear mechanism 53 for shifting, the distance c between the vertical line of the carrier C3 and the vertical line of the sun gear S3 is regarded as 1, and the vertical line of the ring gear R3 is changed from the vertical line of the carrier C3 to the vertical line of the sun gear S3. They are arranged on the opposite side of the line with a distance of c × λ3. For the single pinion type rear planetary gear mechanism 54, the distance d between the vertical line of the sun gear S4 and the vertical line of the ring gear R4 is regarded as 1 + λ4, and the vertical line of the carrier C4 is changed from the vertical line of the sun gear S4 to the vertical line of the ring gear R4. They are arranged on the same side by a distance d / (1 + λ4). The speed diagram shows the first and second rotation generation control brakes B-1 and B-2, the first and second control clutches C-1 and C-2, the first and second control brakes B-3 and B. B-1 to B-4, C-1, and C-2 are entered at points where -4 is selectively activated.
[0029]
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 front ring gear R3 as the first element is connected to the common carriers C1 and C2 of the double planetary gear unit 16 for reduction and is connected to the first control brake B-3, and as the second element. The first-stage carrier C3 and the second-stage ring gear R4 connected to each other are connected in parallel to the second control clutch C-2 and the second control brake B-4, and the second-stage carrier C4 as the third element is connected to the output shaft 18. The front-stage and rear-stage sun gears S3, S4 connected to each other as the fourth element are connected to the first clutch C-1.
[0030]
Hereinafter, the operation of each gear stage will be described. In the case of the first forward shift speed, the first control clutch C-1 and the second control brake B-4 are operated, the input shaft 15 and the front and rear stage sun gears S3 and S4 are connected, and the front stage carrier C3 and the rear stage connected to each other. Since the rotation of the ring gear R4 is restricted, the rotation input to the input shaft 15 is transmitted to the rear stage sun gear S4 via the first control clutch C-1, and the reaction force is supported by the rear stage ring gear R4 whose rotation is restricted. Then, the pinion 34 is revolved and transmitted to the rear carrier C4, and the output shaft 18 is driven to rotate forward at a gear ratio 3.333 of the first gear.
[0031]
In the case of the second forward speed, the first control clutch C-1 and the first control brake B-3 are operated, the input shaft 15 and the front and rear stage sun gears S3 and S4 are connected, and the front stage ring gear R3 is restricted from rotating. The rotation input to the input shaft 15 is transmitted to the preceding stage sun gear S3 via the first control clutch C-1, and the reaction force is supported by the preceding stage ring gear R3 whose rotation is restricted and the pinions 33 are revolved to reciprocate each other. The connected front stage carrier C3 and the rear stage ring gear R4 are rotated, and the pinion 34 is revolved according to the rotational difference between the rear stage sun gear S4 and the rear stage ring gear R4, so that the rear stage carrier C4 and the output shaft 18 have a gear ratio 2 of the second gear. Drive forward at 0.035.
[0032]
In the case of the third forward speed, the first control clutch C-1 and the second rotation generation control brake B-2 are operated, the input shaft 15 and the front and rear stage sun gears S3 and S4 are connected, and the large-diameter ring gear R2 is restricted from rotating. Therefore, the rotation input to the input shaft 15 is transmitted to the front and rear sun gears S3 and S4 via the first control clutch C-1, and to the sun gears S1 and S2 of the double planetary gear unit 16 for reduction. The reaction force is supported by the large-diameter ring gear R2 whose rotation is restricted, and the stepped pinion 25 is revolved to generate the second rotation in the carriers C1 and C2, and the front ring gear R3 of the double planetary gear unit 17 for transmission is The second rotation is transmitted, the pinion 33 is revolved according to the rotational difference between the front stage sun gear S3 and the front stage ring gear R3 to rotate the front stage carrier C3 and the rear stage ring gear R4, and the rear stage sun gear. 4 and is revolving pinion 34 in accordance with the rotational difference between the later stage ring gear R4 to rotate forward driving the output shaft 18 and the subsequent carrier C4 in the gear ratio 1.637 for the third gear position.
[0033]
In the case of the fourth forward speed, the first control clutch C-1 and the first rotation generation control brake B-1 are operated, the input shaft 15 and the front and rear stage sun gears S3 and S4 are connected, and the small-diameter ring gear R1 is restricted from rotating. Therefore, the rotation input to the input shaft 15 is transmitted to the front and rear sun gears S3 and S4 via the first control clutch C-1, and is also transmitted to the sun gears S1 and S2 of the speed reduction double planetary gear unit 16. The reaction force is supported by the small-diameter ring gear R1 whose rotation is restricted, the stepped pinion 25 is revolved to generate the first rotation in the carriers C1 and C2, and the first ring gear R3 of the double planetary gear unit 17 for transmission is The rotation is transmitted, the pinion 33 is revolved according to the rotational difference between the front stage sun gear S3 and the front stage ring gear R3 to rotate the front stage carrier C3 and the rear stage ring gear R4, and the rear stage sun gear. 4 and is revolving pinion 34 in accordance with the rotational difference between the later stage ring gear R4 to rotate forward driving the output shaft 18 and the subsequent carrier C4 in the gear ratio 1.509 for the fourth shift stage.
[0034]
In the case of the fifth forward speed, the first and second control clutches C-1 and C-2 are connected, and the input shaft 15 is connected to the front and rear stage sun gears S3 and S4 and the front stage carrier C3 and the rear stage ring gear R4 connected to each other. Therefore, the rear carrier C4 is also rotated integrally, and the output shaft 18 is driven to rotate forward at the gear ratio 1.000 of the fifth gear.
[0035]
In the case of the sixth forward speed, the second control clutch C-2 and the first rotation generation control brake B-1 are operated, the input shaft 15 is connected to the front carrier C3 and the rear ring gear R4, and the small-diameter ring gear R1 is rotationally restricted. Therefore, the rotation input to the input shaft 15 is transmitted to the front stage carrier C3 and the rear stage ring gear R4 via the second control clutch C-2, and to the sun gears S1 and S2 of the double planetary gear unit 16 for reduction. The reaction force is supported by the small-diameter ring gear R1 that is transmitted and restricted in rotation, and the stepped pinion 25 is revolved to generate the first rotation in the carriers C1 and C2, and the first stage ring gear R3 of the double planetary gear unit 17 for speed change is generated. The first rotation is transmitted, the front and rear sun gears S3 and S4 are rotated according to the rotational difference between the front carrier C3 and the front ring gear R3, and the rear sun gear S4 and the rear phosphorus By revolving the pinion 34 to the forward drive the output shaft 18 and the subsequent carrier C4 in the gear ratio 0.654 sixth gear position according to the rotation difference between the gear R4.
[0036]
In the case of the seventh forward speed, the second control clutch C-2 and the second rotation generation control brake B-2 are operated, the input shaft 15 is connected to the front stage carrier C3 and the rear stage ring gear R4, and the large-diameter ring gear R2 rotates. Therefore, the rotation input to the input shaft 15 is transmitted to the front carrier C3 and the rear ring gear R4 via the second control clutch C-2, and the sun gears S1 and S2 of the double planetary gear unit 16 for speed reduction. , The reaction force is supported by the large-diameter ring gear R2 whose rotation is restricted, the stepped pinion 25 is revolved to generate the second rotation in the carriers C1 and C2, and the front ring gear of the double planetary gear unit 17 for speed change. The second rotation is transmitted to R3, the front and rear sun gears S3 and S4 are rotated according to the rotational difference between the front carrier C3 and the front ring gear R3, and the rear sun gear S4 and the rear ring R By revolving the pinion 34 to the forward drive the output shaft 18 and the subsequent carrier C4 in the gear ratio 0.621 the seventh gear position according to the rotation difference between the gear R4.
[0037]
In the case of the eighth forward speed, the second control clutch C-2 and the first control brake B-3 are operated, the input shaft 15 is connected to the front carrier C3 and the rear ring gear R4, and the front ring gear R3 is restricted from rotating. Therefore, the rotation input to the input shaft 15 is transmitted to the front carrier C3 and the rear ring gear R4 via the second control clutch C-2, and the reaction force is supported by the front ring gear R3 whose rotation is restricted, so that the front and rear stages are supported. The sun gears S3 and S4 are rotated, the pinion 34 is revolved in accordance with the rotational difference between the rear sun gear S4 and the rear ring gear R4, and the rear carrier C4 and the output shaft 18 are driven forward at a gear ratio 0.557 of the eighth gear. To do.
[0038]
In the case of the reverse gear, the first rotation generation control brake B-1 and the second control brake B-4 are operated, and the rotation of the small-diameter ring gear R1, the front carrier C3 and the rear ring gear R4 connected to each other is restricted. The rotation input to the shaft 15 is transmitted to the sun gears S1 and S2, and the reaction force is supported by the small-diameter ring gear R1 whose rotation is restricted, and the stepped pinion 25 is revolved to generate the first rotation on the carriers C1 and C2. Then, the first rotation is transmitted to the front stage ring gear R3 of the double planetary gear unit 17 for speed change, and the front and rear stage sun gears S3 and S4 are reversely rotated via the pinion 33 supported by the front stage carrier C3 whose rotation is restricted. The reaction force is supported by R4, the pinion 34 is revolved, and the rear carrier C4 and the output shaft 18 are driven in reverse at the reverse gear stage gear ratio of 3.726.
[0039]
Rotational ratios of sun gears S1 to S4, carriers C1 to C4, and ring gears R1 to R4 at each gear position when the rotational speed of the sun gears S1 and S2 of the double planetary gear unit 16 for reduction connected to the input shaft 15 is 1. As can be seen from the speed diagram of FIG. 3, the rotation ratio, that is, the gear ratio of the rear carrier C4 at each shift stage is arranged at an appropriate interval, and is separated appropriately according to the automatic transmission according to the present invention. A gear ratio of 8 forward speeds and 1 reverse speed can be obtained.
[0040]
In the first forward shift speed of the first embodiment, the front ring gear R3 is reversely rotated, but the common carriers C1 and C2 directly connected to the front ring gear R3 are also reversely rotated. At this time, since the common sun gears S1 and S2 rotate at the same rotational speed as the input shaft 15, the small-diameter ring gear R1 and the large-diameter ring gear R2 rotate in the reverse direction larger than the common carriers C1 and C2. Then, a large relative rotational difference occurs between the common sun gears S1 and S2 and the small diameter and large diameter ring gears R1 and R2, and as a result, the small diameter pinion 23 and the large diameter pinion 24 meshing with both rotate at high speed.
[0041]
Therefore, in order to prevent this high speed rotation, it is newly added either between the carriers C1 and C2 of the first embodiment and the front ring gear R3, or between the common sun gears S1 and S2 and the input shaft 15. It is also possible to selectively connect with an intermediate clutch.
[0042]
In this case, the new clutch is engaged at the shift speed at which power is transmitted from the carriers C1 and C2 to the front ring gear R3, that is, the forward third, fourth, sixth and seventh shift speeds and the reverse shift speed. Although it can be released at the forward first, second, fifth, and eighth shift stages, which are the shift stages, it may be opened only at the forward first shift stage where the upper high-speed rotation is remarkable.
[0043]
Next, a second embodiment will be described based on FIG. In the second embodiment, the shift type planetary gear unit 17, the first and second control clutches C-1 and C-2, the first and second control brakes B-3 and B-4, Since it is the same as that of the embodiment, the same reference numerals are attached to the drawings and description thereof is omitted, and only the reduction planetary gear device 60 different from the first embodiment will be described.
[0044]
As shown in FIG. 4, the reduction type planetary gear unit 60 is configured by connecting and sharing the sun gears S <b> 1 and S <b> 2 and the carriers C <b> 1 and C <b> 2 of the single pinion type planetary gear mechanism 61 and the double pinion type planetary gear mechanism 62. It is configured. That is, the common sun gears S1 and S2 that are rotatably supported on the common axis 13, the first ring gear R1 that is rotatably supported on the common axis 13 and meshes with the sun gears S1 and S2 via the long pinion 63, the common axis The second ring gear R2, the long pinion 63 and the intermediate pinion 64, which are rotatably supported on the gear 13 and meshed with the sun gears S1 and S2 via the long pinion 63 and the intermediate pinion 64, are rotatably supported on the common axis 13. It is composed of supported common carriers C1 and C2. The common sun gears S1 and S2 are connected to the input shaft 15, and the second ring gear R2 is connected to the front ring gear R3 of the double planetary gear unit 17 for transmission.
[0045]
In the first ring gear R1, the first ring gear R1 is selectively connected to the transmission case 12, the rotation is restricted, and the second rotation is generated in the second ring gear R2 with a rotation speed smaller than the rotation of the input shaft 15. The generation control brake B-1 is coupled, and the carriers C1 and C2 are selectively connected to the transmission case 12 to the common carriers C1 and C2, and the rotation is restricted to the second ring gear R2 from the first rotation. A second rotation generation control brake B-2 for generating a small second rotation is connected.
[0046]
The reduction type planetary gear unit 60 includes a gear reduction unit 49 that is connected to the input shaft 15 and generates a first rotation having a rotation speed smaller than the rotation of the input shaft 15 and a second rotation having a rotation speed smaller than the first rotation. Constitute. Then, the first and second rotation generation control brakes B-1 and B-2 select one of the first and second rotations generated in the second ring gear R2 of the double planetary gear device for deceleration 60 and perform the shift. The rotation selection means 50 which transmits to the front | former stage ring gear R3 of the double planetary gear apparatus 17 is comprised.
[0047]
Also in the second embodiment, the input shaft 15 is rotated by the first and second control clutches C-1 and C-2 so that the front and rear sun gears S3 and S4 of the shift type planetary gear unit 17 are connected to each other and the front carrier C3 connected to each other. Are selectively transmitted to the rear ring gear R4, and the first or second rotation generated in the second ring gear R2 is selectively transmitted to the front ring gear R3 so that the rotation of the input shaft 15 is 8 forward stages and 1 reverse stage. Since shifting to the same speed as in the case of the first embodiment is not described in detail. FIG. 5 shows the operating state of each control clutch and control brake at each speed stage.
[0048]
FIG. 5 also shows that the gear ratio λ1 of the first planetary gear mechanism 61 for reduction comprising the sun gear S1, long pinion 63, carrier C1 and ring gear R1 of the double planetary gear device 60 for reduction is 0.508, sun gear S2, long pinion 63, the gear ratio λ2 of the second planetary gear mechanism 62 for reduction comprising the intermediate pinion 64, the carrier C2 and the ring gear R2 is 0.341, the front stage sun gear S3, the pinion 33, the front stage carrier C3 and the front stage of the double planetary gear unit 17 for transmission. The gear ratio λ3 of the transmission front planetary gear mechanism 53 including the ring gear R3 is 0.325, and the gear ratio λ4 of the transmission rear planetary gear mechanism 54 including the rear sun gear S4, the pinion 34, the rear carrier C4, and the rear ring gear R4 is 0. 381, the gear ratio at each shift speed (the rotational speed of the input shaft 15 / the output shaft 18). Number rolling) is shown on the gear ratio column.
[0049]
In the double pinion type second planetary gear mechanism 62 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 expressed by the following equation (2). Indicated. The gear ratio at each gear stage is calculated based on the equation (2) and the relational equation (1) in the single pinion planetary gear mechanism described above. Also in the single pinion type planetary gear mechanism, assuming that the number of teeth of the sun gear S2 is Zs2 and the number of teeth of the second ring gear R2 is Zr2, the gear ratio of the second planetary gear mechanism 62 for reduction is λ2 = Zs2 / Zr2.
Nr = (1-λ) Nc + λNs (2)
[0050]
The first and second control clutches C-1 and C-2 are selectively connected, and the first and second control brakes B-3 and B-4 are selectively operated to control the first and second rotation generation control. When the brakes B-1 and B-2 are selectively operated, the speed ratios of the elements of the speed reduction double planetary gear device 60 and the speed change double planetary gear device 17 are as shown in the speed diagram of FIG. . Regarding the double pinion type second planetary gear mechanism 62 for reduction, the distance b 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 changed from the vertical line of the carrier C2 to the sun gear S2. They are arranged on the same side as the vertical line by a distance of b × λ2. The speed diagram shows the first and second rotation generation control brakes B-1 and B-2, the first and second control clutches C-1 and C-2, the first and second control brakes B-3 and B. B-1 to B-4, C-1, and C-2 are entered at points where -4 is selectively activated. As shown in the velocity diagram of FIG. 6, the front ring gear R3 of the shift type planetary gear unit 17 is the first element, the front carrier C3 and the rear ring gear R4 connected to each other are the second element, and the rear carrier C4 is the third element. The front and rear sun gears S3 and S4 connected to each other are the fourth element. Since the operating states of the control clutch and the control brake at each speed stage of the double planetary gear unit 17 for shifting are the same as those in the first embodiment shown in FIG. 2, detailed description of the operation is omitted.
[0051]
In the second embodiment, for the same reason as in the first embodiment, either between the second ring gear R2 and the front ring gear R3 or between the input shaft 15 and the common sun gears S1 and S2. A new clutch can be interposed between the two.
[0052]
Next, a third embodiment in which the gear reduction device 49 is constituted by a reduction gear train will be described. 3rd Embodiment is 1st Embodiment about the double planetary gear apparatus 17 for transmission, 1st, 2nd control clutch C-1, C-2, 1st, 2nd control brake B-3, B-4. 7, the same reference numerals are attached to FIG. 7, and the description thereof is omitted. The speed reduction gear train 65 and the connection relationship between the speed reduction gear train 65 and the speed change double planetary gear device 17 are different from those of the first embodiment. To do. A turbine 47 of the fluid torque converter 11 is connected to an input shaft 66 rotatably supported by the transmission case 12 of the automatic transmission 10, and large diameter, medium diameter, and small diameter gears 67, 68, 69 are fixed to the input shaft 66. Has been. A first gear 70 having the same diameter that meshes with the large-diameter gear 67 is rotatably supported on the axis 13 of the double planetary gear device 17 for transmission, and meshes with the medium-diameter and small-diameter gears 68 and 69, respectively. Gears 71 and 72 are rotatably supported on the axis 13. As a result, the first gear 70 rotates at the same rotational speed as the input shaft 66, the second gear 71 rotates at a first rotational speed that is smaller than the rotational speed of the input shaft 66, and the third gear 72 rotates at the first rotational speed. It rotates in the second rotation with a smaller rotational speed. The first gear 70 is detachably connected to the front and rear sun gears S1, S2 of the shift type planetary gear unit 17 by the first control clutch C-1, and is connected to the front carrier C3 and the rear ring gear by the second control clutch C-2. It is detachably connected to R4. The second and third gears 71 and 72 are detachably connected to the front ring gear R3 by the first and second rotation transmission control clutches C-4 and C-3. The first and second rotation transmission control clutches C-4 and C-3 select one of the first and second rotations generated by the reduction gear train 65 to select the front stage ring gear R3 of the double planetary gear unit 17 for shifting. Rotation selection means 50 for transmitting to is configured.
[0053]
As shown in FIG. 8, the speed diagram of the shift type planetary gear unit 17 is the same as that of the first embodiment, the front ring gear R3 is the first element, the front carrier C3 and the rear ring gear R4 are the second element, The rear carrier C4 is the third element, and the front and rear sun gears S3 and S4 are the fourth element. The operating state of the control clutch and the control brake at each speed stage of the shift type planetary gear unit 17 is changed to the first and second rotation transmission control clutches instead of the first and second rotation generation control brakes B-1 and B-2. If C-4 and C-3 are operated, the operation is the same as that of the first embodiment shown in FIG.
[0054]
Next, another embodiment in which the same speed reduction double planetary gear device 16 as that of the first embodiment is used for the gear reduction device 49, and the speed change double planetary gear device is constituted by a double planetary gear device different from the above embodiment. explain. Since the double planetary gear unit 16 for speed reduction is the same as that of the first embodiment, the same reference numerals are given to the drawings and the description thereof is omitted.
[0055]
In the fourth embodiment, as shown in FIG. 9, the speed-changing double planetary gear device 75 includes two sets of front and rear stages of single pinion type planetary gear mechanisms 76 and 77, and a front stage carrier C 3 that supports the pinion 78. And the rear ring gear R4 are connected, and the rear carrier C4 for supporting the pinion 79 and the front ring gear R3 are connected. An output shaft 18 is connected to the rear carrier C4. The front stage sun gear S3 is connected to the carriers C1 and C2 of the double planetary gear unit 16 for reduction, and the rear stage sun gear S4 is detachably connected to the input shaft 15 via the first control clutch C-1, and is connected to the front stage carrier C3 and the rear stage. The ring gear R4 is detachably connected to the input shaft 15 via the second control clutch C-2. The front stage sun gear S3 is connected to the first control brake B-3, and the front stage carrier C3 and the rear stage ring gear R4 are connected to the second control brake B-4 to selectively restrict the rotation.
[0056]
In the fourth embodiment, for the same reason as in the first embodiment, either between the carriers C1 and C2 and the preceding sun gear S3 or between the input shaft 15 and the common sun gears S1 and S2. A new clutch can be interposed between the two.
[0057]
As shown in the speed diagram of FIG. 10, the front stage sun gear S3 is the first element, the front stage carrier C3 and the rear stage ring gear R4 connected to each other is the second element, and the front stage ring gear R3 and the rear stage carrier C4 connected to each other are the third element. The rear stage sun gear S4 is the fourth element. Since the operation states of the control clutch and the control brake at each speed stage are the same as those in the first embodiment shown in FIG. 2, detailed description of the operation is omitted.
[0058]
In the fifth embodiment, as shown in FIG. 11, the double planetary gear device 80 for speed change includes a front stage and rear stage carriers C3 and C4 of a single pinion type front planetary gear mechanism 81 and a double pinion type rear planetary gear mechanism 82. The front and rear ring gears R3 and R4 are connected and shared, respectively. That is, a long pinion 84, a long pinion 84 and an intermediate gear which are directly meshed with the front and rear sun gears S3 and S4 and the front sun gear S3 rotatably supported on the common axis 13 and meshed with the rear sun gear S4 via the intermediate pinion 83. The front and rear stage gears R3 that mesh with the front and rear carriers C3 and C4 and the long pinion 84 that are rotatably supported on the common axis 13 and rotatably supported on the common axis 13 are supported. , R4. The ring gears R3 and R4 are connected to the output shaft 18. The front-stage sun gear S3 is connected to the carriers C1 and C2 of the double planetary gear unit 16 for reduction, and the rear-stage sun gear S4 is detachably connected to the input shaft 15 via the first control clutch C-1. C3 and C4 are detachably connected to the input shaft 15 via the second control clutch C-2. The front stage sun gear S3 is connected to the first control brake B-3, and the front and rear stage carriers C3 and C4 are connected to the second control brake B-4 to selectively restrict the rotation.
[0059]
In the fifth embodiment, for the same reason as in the first embodiment, either between the carriers C1 and C2 and the preceding sun gear S3 or between the input shaft 15 and the common sun gears S1 and S2. A new clutch can be interposed between the two.
[0060]
As shown in the velocity diagram of FIG. 12, the front stage sun gear S3 is the first element, the front stage and rear stage carriers C3 and C4 connected to each other are the second element, and the front stage and rear stage ring gears R3 and R4 connected to each other are the third element. The rear stage sun gear S4 is the fourth element. Since the operating state of the control clutch and the control brake at each speed stage of the shift type planetary gear device 80 is the same as that of the first embodiment shown in FIG. 2, detailed description of the operation is omitted.
[0061]
In the sixth embodiment, as shown in FIG. 13, the double planetary gear device 90 for shifting includes a front stage and rear stage sun gears S3 and S4 of a single pinion type front planetary gear mechanism 91 and a double pinion type rear planetary gear mechanism 92. The front and rear carriers C3 and C4 are connected and shared, respectively. That is, the common sun gears S3 and S4 that are rotatably supported on the common axis 13, the ring gear R3 that is rotatably supported on the common axis 13 and meshes with the sun gears S3 and S4 via the long pinion 93, and the common axis 13 The ring gear R4, the long pinion 93 and the intermediate pinion 94 which are rotatably supported by the sun gears S3 and S4 and meshed with the long pinion 93 and the intermediate pinion 94 are supported on the common axis 13 so as to be rotatably supported. Carrier C3, C4. An output shaft 18 is connected to the ring gear R4. The front ring gear R3 is connected to the carriers C1 and C2 of the double planetary gear unit 16 for reduction, and the front and rear sun gears S3 and S4 are detachably connected to the input shaft 15 via the first control clutch C-1. The front and rear carriers C3 and C4 are detachably connected to the input shaft 15 via the second control clutch C-2. The front ring gear R3 is connected to the first control brake B-3, and the front and rear stage carriers C3 and C4 are connected to the second control brake B-4 to selectively restrict the rotation.
[0062]
In the sixth embodiment, for the same reason as in the first embodiment, either between the carriers C1 and C2 and the front ring gear R3 or between the input shaft 15 and the common sun gears S1 and S2. A new clutch can be interposed between the two.
[0063]
As shown in the velocity diagram of FIG. 14, the front ring gear R3 is the first element, the front and rear carriers C3 and C4 connected to each other is the second element, the rear ring gear R4 is the third element, the front and rear stages connected to each other. The sun gears S3 and S4 are the fourth element. Since the operation states of the control clutch and the control brake at each speed stage are the same as those in the first embodiment shown in FIG. 2, detailed description of the operation is omitted.
[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.
FIG. 7 is a skeleton diagram showing a third embodiment.
FIG. 8 is a velocity diagram showing a rotation ratio of each element of the shift type planetary gear device in each speed stage of the third embodiment.
FIG. 9 is a skeleton diagram showing a fourth embodiment.
FIG. 10 is a velocity diagram showing a rotation ratio of each element of the planetary gear device at each speed stage of the fourth embodiment.
FIG. 11 is a skeleton diagram showing a fifth embodiment.
FIG. 12 is a velocity diagram showing a rotation ratio of each element of the planetary gear device at each speed stage of the fifth embodiment.
FIG. 13 is a skeleton diagram showing a sixth embodiment.
FIG. 14 is a velocity diagram showing a rotation ratio of each element of the planetary gear device at each speed stage of the sixth embodiment.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 10 ... Automatic transmission, 11 ... Fluid torque converter, 12 ... Transmission case, 13 ... Common axis, 15, 66 ... Input shaft, 16, 60 ... Double planetary gear for reduction Device, 17, 75, 80, 90... Double planetary gear device for transmission, 18... Output shaft, 23... Small diameter pinion, 24... Large diameter pinion, 25. , 34, 78, 79... Pinion, 64, 83, 94... Intermediate pinion, 63, 84, 93... Long pinion, 49. , 52, 53, 54, 61, 76, 77, 81, 91... Single pinion type planetary gear mechanism, 62, 82, 92... Double pinion type planetary gear mechanism, 65. Column, S1, S2, S3, S ... Sun gear, C1, C2, C3, C4 ... Carrier, R1, R2, R3, R4 ... Ring gear, C-1, C-2 ... First, second control clutch, C-3 , C-4 ... second, first rotation transmission control clutch, B-1, B-2 ... first, second rotation generation control brake, B-3, B-4 ... first, Second control brake.

Claims (8)

入力軸と、該入力軸に連結され該入力軸の回転より回転数が小さい第１回転及び該第１回転より回転数が小さい第２回転を生成する歯車減速装置と、速度線図においてギヤ比に対応した間隔で順次並べられた４個の要素に並び順にそれぞれ対応する第１、第２、第３及び第４要素を有する変速用複式遊星歯車装置と、前記入力軸の回転を第２、第４要素に選択的に伝達する第２、第１制御クラッチと、前記第１及び第２回転の一方を選択して前記第１要素に伝達する回転選択手段と、前記第１、第２要素の回転を選択的に規制する第１及び第２制御ブレーキと、前記第３要素に連結された出力軸とを備え、
前記変速用複式遊星歯車装置は、
前記歯車減速装置から前記第１回転及び前記第２回転の一方が前記回転選択手段によって選択されて前記第１要素に伝達された場合に、前記入力軸の回転が前記第１制御クラッチで前記第４要素に伝達されることにより前記入力軸の回転数より前記出力軸の回転数が小さい減速段を形成し、
前記歯車減速装置から前記第１回転及び前記第２回転の一方が前記回転選択手段によって選択されて前記第１要素に伝達された場合に、前記入力軸の回転が前記第２制御クラッチで前記第２要素に伝達されることにより前記入力軸の回転数より前記出力軸の回転数が大きい増速段を形成することを特徴とする自動変速機。An input shaft, a gear reduction device connected to the input shaft and generating a first rotation having a rotation speed smaller than the rotation of the input shaft and a second rotation having a rotation speed smaller than the first rotation; and a gear ratio in the speed diagram A dual planetary gear device for transmission having first, second, third and fourth elements respectively corresponding to the four elements sequentially arranged at intervals corresponding to the second, and second rotation of the input shaft. Second and first control clutches for selectively transmitting to the fourth element; rotation selecting means for selecting one of the first and second rotations and transmitting to the first element; and the first and second elements First and second control brakes for selectively restricting rotation of the motor, and an output shaft connected to the third element,
The shift type planetary gear unit for shifting is
If one of the first rotation and the second rotation from the gear reduction device is transmitted to the first element is selected by the rotary selection means, the rotation of the input shaft is in said first control clutch first By transmitting to four elements, a speed reduction stage in which the rotational speed of the output shaft is smaller than the rotational speed of the input shaft is formed,
When one of the first rotation and the second rotation is selected from the gear reduction device by the rotation selection means and transmitted to the first element, the rotation of the input shaft is performed by the second control clutch. An automatic transmission having a speed increasing stage in which the rotational speed of the output shaft is larger than the rotational speed of the input shaft by being transmitted to two elements. 請求項１に記載の自動変速機において、入力軸に連結されたサンギヤ、該サンギヤと噛合する小径ピニオン及び大径ピニオンからなる段付ピニオンを支承するキャリヤ、小径ピニオンと噛合する小径リングギヤ並びに前記大径ピニオンと噛合する大径リングギヤからなる減速用複式遊星歯車装置で前記歯車減速装置を構成し、該減速用複式遊星歯車装置のキャリヤを前記変速用複式遊星歯車装置の第１要素に連結し、前記小径及び大径リングギヤの回転を夫々選択的に規制して前記キャリヤに前記第１、第２回転を選択的に生成させる第１、第２回転生成制御ブレーキで前記回転選択手段を構成したことを特徴とする自動変速機。 2. The automatic transmission according to claim 1, wherein a sun gear connected to an input shaft, a carrier supporting a stepped pinion comprising a small-diameter pinion and a large-diameter pinion meshing with the sun gear, a small-diameter ring gear meshing with a small-diameter pinion, and the large-diameter The reduction gear planetary gear device comprising a large-diameter ring gear meshing with a diameter pinion constitutes the gear reduction device, and the carrier of the reduction double planetary gear device is connected to the first element of the transmission double planetary gear device; The rotation selection means is configured by first and second rotation generation control brakes that selectively restrict the rotation of the small-diameter and large-diameter ring gears and cause the carrier to selectively generate the first and second rotations. Automatic transmission characterized by. 請求項１に記載の自動変速機において、入力軸に連結されたサンギヤ、該サンギヤと噛合するロングピニオン及び該ロングピニオンと噛合する中間ピニオンを支承するキャリヤ、前記ロングピニオンと噛合する第１リングギヤ並びに前記中間ピニオンと噛合する第２リングギヤからなる減速用複式遊星歯車装置で前記歯車減速装置を構成し、該減速用複式遊星歯車装置の第２リングギヤを前記変速用複式遊星歯車装置の第１要素に連結し、前記第１リングギヤ及びキャリヤの回転を夫々選択的に規制して前記第２リングギヤに前記第１、第２回転を選択的に生成させる第１、第２回転生成制御ブレーキで前記回転選択手段を構成したことを特徴とする自動変速機。 2. The automatic transmission according to claim 1, wherein a sun gear coupled to the input shaft, a long pinion meshing with the sun gear, a carrier supporting an intermediate pinion meshing with the long pinion, a first ring gear meshing with the long pinion, and The gear reduction device is constituted by a reduction double planetary gear device comprising a second ring gear meshing with the intermediate pinion, and the second ring gear of the reduction double planetary gear device is used as a first element of the transmission double planetary gear device. The rotation selection is performed by first and second rotation generation control brakes that are coupled and selectively restrict the rotation of the first ring gear and the carrier, respectively, and cause the second ring gear to selectively generate the first and second rotations. An automatic transmission characterized by comprising means. 請求項１に記載の自動変速機において、入力軸に固定された大径、中径、小径歯車、前記変速用遊星歯車装置と同心に回転可能に支承され前記大径、中径、小径歯車と夫々噛合して入力回転、該入力回転より回転数が小さい第１回転、該第１回転より回転数が小さい第２回転を夫々生成する入力歯車、第１歯車、第２歯車からなる減速用歯車列で前記歯車減速装置を構成し、前記入力歯車を前記第２、第１制御クラッチに連結して前記入力回転を前記変速用複式遊星歯車装置の第２、第４要素に選択的に伝達し、前記第１、第２歯車を前記第１要素に夫々係脱可能に連結する第１、第２回転伝達制御クラッチにより前記第１及び第２回転の一方を選択して前記第１要素に伝達する回転選択手段を構成したことを特徴とする自動変速機。 2. The automatic transmission according to claim 1, wherein the large, medium, and small diameter gears fixed to the input shaft are rotatably supported concentrically with the planetary gear device for transmission, and the large, medium, and small diameter gears A reduction gear composed of an input gear, a first gear, and a second gear, each of which meshes with each other to generate an input rotation, a first rotation having a rotation speed smaller than the input rotation, and a second rotation having a rotation speed smaller than the first rotation. The gear reduction device is configured in a row, the input gear is connected to the second and first control clutches, and the input rotation is selectively transmitted to the second and fourth elements of the shift type planetary gear device. One of the first and second rotations is selected and transmitted to the first element by first and second rotation transmission control clutches that detachably connect the first and second gears to the first element, respectively. An automatic transmission characterized by comprising a rotation selecting means. 請求項１乃至請求項４のいずれかに記載の自動変速機において、前記変速用複式遊星歯車装置を前後段２組のシングルピニオン型の遊星歯車機構とし、前段リングギヤを第１要素とし、前段キャリアと後段リングギヤとを連結して第２要素とし、後段キャリヤを第３要素とし、前段及び後段サンギヤを連結して第４要素としたことを特徴とする自動変速機。 5. The automatic transmission according to claim 1, wherein the double planetary gear device for shifting is a single-pinion type planetary gear mechanism having two sets of front and rear stages, a front ring gear is a first element, and a front carrier And the rear ring gear are connected as a second element, the rear carrier is a third element, and the front and rear sun gears are connected as a fourth element. 請求項１乃至請求項４のいずれかに記載の自動変速機において、前記変速用複式遊星歯車装置を前後段２組のシングルピニオン型の遊星歯車機構とし、前段サンギヤを第１要素とし、前段キャリヤと後段リングギヤとを連結して第２要素とし、前段リングギヤと後段キャリヤとを連結して第３要素とし、後段サンギヤを第４要素としたことを特徴とする自動変速機。 5. The automatic transmission according to claim 1, wherein the double planetary gear device for shifting is a single-pinion type planetary gear mechanism having two sets of front and rear stages, the front sun gear is a first element, and the front carrier And the rear ring gear are connected as a second element, the front ring gear and the rear carrier are connected as a third element, and the rear sun gear is a fourth element. 請求項１乃至請求項４のいずれかに記載の自動変速機において、前記変速用複式遊星歯車装置をシングルピニオン型の前段遊星歯車機構とダブルピニオン型の後段遊星歯車機構とし、前段サンギヤを第１要素とし、前段及び後段キャリヤを連結して第２要素とし、前段及び後段リングギヤを連結して第３要素とし、後段サンギヤを第４要素としたことを特徴とする自動変速機。 5. The automatic transmission according to claim 1, wherein the double planetary gear device for shifting is a single-pinion type front planetary gear mechanism and a double-pinion type rear planetary gear mechanism, and the front-stage sun gear is a first-stage sun gear. An automatic transmission comprising: an element, a front stage and a rear stage carrier connected to form a second element; a front stage and a rear stage ring gear connected to a third element; and the rear stage sun gear to a fourth element. 請求項１乃至請求項４のいずれかに記載の自動変速機において、前記変速用複式遊星歯車装置をシングルピニオン型の前段遊星歯車機構とダブルピニオン型の後段遊星歯車機構とし、前段リングギヤを第１要素とし、前段及び後段キャリヤを連結して第２要素とし、後段リングギヤを第３要素とし、前段及び後段サンギヤを連結して第４要素としたことを特徴とする自動変速機。 5. The automatic transmission according to claim 1, wherein the double planetary gear device for shifting is a single-pinion type front planetary gear mechanism and a double-pinion type rear planetary gear mechanism, and the front-stage ring gear is a first-stage ring gear. An automatic transmission comprising: an element, a front stage and a rear stage carrier coupled to form a second element; a rear stage ring gear serving as a third element; and a front stage and rear stage sun gear coupled to a fourth element.
JP2001011855A 2001-01-19 2001-01-19 Automatic transmission Expired - Fee Related JP4590742B2 (en)

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