JP4019764B2

JP4019764B2 - Control device for automatic transmission

Info

Publication number: JP4019764B2
Application number: JP2002095415A
Authority: JP
Inventors: 信一朗村上; 修司森; 正宏早渕; 正明西田; 洋筒井
Original assignee: Aisin AW Co Ltd
Current assignee: Aisin AW Co Ltd
Priority date: 2002-03-29
Filing date: 2002-03-29
Publication date: 2007-12-12
Anticipated expiration: 2022-03-29
Also published as: JP2003287122A

Description

【０００１】
【発明の属する技術分野】
本発明は、自動変速機のマニュアルバルブを走行中にニュートラルレンジにシフトし、前進走行レンジに戻したとき、エンジンが過回転数で回転されることを防止する自動変速機の制御装置に関するものである。
【０００２】
【従来の技術】
車両の走行中にシフトレバーが操作されてマニュアルバルブが前進走行レンジからニュートラルレンジにシフトされ、前進走行レンジに戻されることがある。従来、マニュアルバルブをニュートラルレンジから前進走行レンジに切換えたときに行われる油圧漸増制御においては、第５又は第６速段を成立するために第２クラッチを係合するとき、係合ショックを低減するために、第２クラッチの油圧駆動部に供給される油圧サーボ装置の出力油圧を漸増していた。
【０００３】
【発明が解決しようとする課題】
ところが、低速段の成立時に常時係合される第１クラッチに油圧を給排する油圧サーボ装置が、リニアソレノイドをオン状態にしても出力油圧を送出し第１クラッチを異常係合する故障のときに、例えば第５速段で走行中にマニュアルバルブをニュートラルレンジＮにシフトすると、図５において、マニュアルバルブ２５のポートＤから第１、第２クラッチC-1，C-2用の油圧サーボ装置２６への油圧の供給が遮断され、第１、第２クラッチC-1，C-2が解放され、ソレノイドSLC-3，SLB-1がオン状態にされて第３クラッチC-3、第１ブレーキB-1用の油圧サーボ装置２６からの出力油圧がゼロになり第３クラッチC-1、第１ブレーキB-1が解放されてニュートラルとなる。制御ポートに油圧が供給されなくなったカットオフ弁５１は第２クラッチC-2用の油圧サーボ装置２６をポートＤから遮断する。その後にマニュアルバルブ２５が前進走行レンジＤにシフトされ、第５速段を指令して油圧漸増制御が行なわれると、ソレノイドSLC-3がオフ状態にされて第３クラッチC-3が係合され、ポートＤから第１クラッチC-1用の油圧サーボ装置２６に油圧が供給され、第１クラッチC-1が異常係合されて第３速段が成立する。ソレノイド弁４８が所定時間オン状態にされてカットオフ弁５１がシフトされ、油圧がポートＤから第２クラッチC-2用の油圧サーボ装置２６に供給され、ソレノイドSLC-2に印加される制御電流が漸減されてオフ状態にされ、第2クラッチC-2が少し遅れて係合される。制御ポートに第２、第３クラッチC-2，C-3に供給される油圧が供給されてカットオフ弁５０がシフトされ、第１クラッチC-1用の油圧サーボ装置２６がポートDから遮断され、第１クラッチC-1が解放される。
【０００４】
このように、第１クラッチC-1が異常係合する故障のときに、第５速段を指令して油圧漸増制御を行なうと、制御の途中で第３速段が成立し、エンジンが駆動輪によって連れ回りされて過回転数で回転される可能性があった。同様に、第６速段を指令して油圧漸増制御を行うと、制御の途中で第２速段が成立する。
【０００５】
本発明は、係る従来の不具合を解消するためになされたもので、自動変速装置の低速段成立時に常時係合される係合要素が異常係合する故障のときに、油圧漸増制御してもエンジンが駆動輪によって過回転数で連れ回りされることがないようにすることである。
【０００６】
【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に係る発明の構成上の特徴は、入力軸の回転を変速して出力軸に出力する変速機構と、該変速機構の各変速段を成立するために係脱される複数の係合要素と、ソレノイドに供給される電気信号に応じた出力油圧を給排して前記係合要素を夫々係脱させる複数の油圧サーボ装置と、前記変速機構の入力軸の回転数及び出力軸の回転数を夫々検出する入力軸回転数検出装置及び出力軸回転数検出装置を備えた自動変速機の制御装置において、前記複数の係合要素は、前記変速機構の高速段成立時に常時係合される高速段係合要素と、低速段成立時に常時係合される低速段係合要素と、前記高速段係合要素又は前記低速段係合要素と共に一つが係合されて前記高速段又は低速段の各変速段を成立させる第３係合要素及び第４係合要素を含み、前記高速段を成立させるために前記高速段係合要素及び前記第３係合要素又は第４係合要素が係合することにより、前記低速段係合要素への油圧の供給を遮断する遮断手段を有し、前記制御装置は、走行中に前記変速機構をニュートラルから高速段のいずれかの変速段に切換える指令が送出されたとき、前記出力軸の回転数が所定低速回転数より小さいか否か判定する手段と、小さい場合は、前記油圧サーボ装置から前記高速段係合要素に供給される油圧を漸増して前記変速機構をニュートラルから前記指令された変速段に切換える油圧漸増制御を行う手段と、大きい場合は、前記油圧漸増制御を中断し又は行わずに前記高速段係合要素に油圧を給排する油圧サーボ装置の出力油圧を急峻に高くして前記高速段係合要素を早急に係合する油圧急峻制御を行う手段を備えたことである。
【０００７】
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の制御装置において、前記所定低速回転数は、前記低速段係合要素が異常係合されているときに、前記油圧漸増制御の途中で成立する低速段のギヤ比を乗じるとエンジンの制限回転数と等しくなる出力軸の回転数であることである。
【０００８】
請求項３に係る発明の構成上の特徴は、請求項１又は２に記載の自動変速機の制御装置において、前記制御装置は、前記出力軸の回転数が前記所定低速回転数より大きい場合は、前記入力軸の回転数が前記出力軸の回転数に前記指令された変速段のギヤ比を乗じた理想回転数より大きいか否か、又は前記入力軸の回転数と出力軸の回転数との実際のギヤ比が前記指令された変速段のギヤ比より大きいか否か判定する手段と、大きい場合は、エンジンのトルクリミテーション制御を行って前記入力軸の回転数が前記理想回転数とほぼ等しくなったとき、前記油圧漸増制御を行わずに前記高速段係合要素に油圧を給排する油圧サーボ装置の出力油圧を急峻に高くして前記高速段係合要素を早急に係合する油圧急峻制御を行う手段を備えたことである。
【０００９】
請求項４に係る発明の構成上の特徴は、請求項３に記載の自動変速機の制御装置において、前記制御装置は、前記出力軸の回転数が前記所定低速回転数より大きい場合は、前記入力軸の回転数又は実際のギヤ比が前記理想回転数又は前記指令された変速段のギヤ比より小さい場合、前記入力軸の回転数が前記理想回転数とほぼ等しくなったとき、前記油圧漸増制御を中断して前記高速段係合要素に油圧を給排する油圧サーボ装置の出力油圧を急峻に高くして前記高速段係合要素を早急に係合する油圧急峻制御を行う手段を備えたことである。
【００１０】
請求項５に係る発明の構成上の特徴は、請求項１乃至４のいずれかに記載の自動変速機の制御装置において、前記制御装置は、前記出力軸の回転数が所定高速回転数より大きいか否か判定する手段と、大きい場合は、前記出力軸の回転数が該所定高速回転数以下に低下するまで待機する手段を備えたことである。
【００１１】
請求項６に係る発明の構成上の特徴は、請求項５に記載の自動変速機の制御装置において、前記所定高速回転数は、ニュートラルから指令された高速段の変速段に切換える油圧漸増制御を行った場合に成立する可能性がある変速段のギヤ比又は指令された変速段のギヤ比を乗じるとエンジンの制限回転数と等しくなる出力軸の回転数であることである。
【００１２】
請求項７に係る発明の構成上の特徴は、請求項１乃至６のいずれかに記載の自動変速機の制御装置において、前記高速段係合要素用の油圧サーボ装置に油圧が印加され前記高速段係合要素が係合されると、前記低速段係合要素用の油圧サーボ装置に印加される油圧を遮断するカットオフ弁を有することである。
【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、走行中に変速機構をニュートラルから高速段のいずれかの変速段に切換える指令が送出されたとき、エンジンが駆動輪によって過回転数で連れ回りされることを防止するために、出力軸の回転数が所定低速回転数より小さいか否か判定し、小さい場合は、油圧サーボ装置から高速段係合要素に供給される油圧を漸増して変速機構をニュートラルから指令された変速段に切換える油圧漸増制御を行う。大きい場合は、油圧漸増制御を中断し又は行わずに高速段係合要素を早急に係合する油圧急峻制御を行う。これにより、例えば走行中にマニュアルバルブがニュートラルレンジＮにシフトされた後に、前進走行レンジDに戻されても、エンジンが駆動輪によって過回転数で連れ回りされることがない。
【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、低速段係合要素が異常係合されているときに、油圧漸増制御の途中で成立する低速段のギヤ比を乗じるとエンジンの制限回転数と等しくなる所定低速回転数より出力軸の回転数が小さい場合は、油圧漸増制御を行い、大きい場合は、高速段係合要素を早急に係合するようにしたので、エンジンが駆動輪によって過回転数で連れ回りされることを確実に防止することができる。
【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、入力軸の回転数が前記出力軸の回転数に指令された変速段のギヤ比を乗じた理想回転数より大きい場合、換言すれば入力軸の回転数と出力軸の回転数との実際のギヤ比が指令された変速段のギヤ比より大きい場合は、エンジンのトルクリミテーション制御を行って入力軸の回転数が理想回転数とほぼ等しくなったとき、油圧漸増制御を行わずに高速段係合要素に油圧を給排する油圧サーボ装置の出力油圧を急峻に高くして高速段係合要素を早急に係合する油圧急峻制御を行うので、油圧漸増制御の途中で低速段が成立してもエンジンが駆動輪によって過回転数で連れ回りされることがない。
【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、入力軸の回転数が出力軸の回転数に指令された変速段のギヤ比を乗じた理想回転数より小さい場合、入力軸の回転数が理想回転数とほぼ等しくなったとき、油圧漸増制御を中断して高速段係合要素に油圧を給排する油圧サーボ装置の出力油圧を急峻に高くして高速段係合要素を早急に係合する油圧急峻制御を行うので、油圧漸増制御の途中で低速段が成立してもエンジンが駆動輪によって過回転数で連れ回りされることがない。
【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、出力軸の回転数が所定高速回転数より大きい場合は、出力軸の回転数が該所定高速回転数以下に低下するまで待機するので、エンジンが駆動輪によって過回転数で連れ回りされることを確実に防止することができる。
【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、所定高速回転数より出力軸の回転数が大きい場合は、所定高速回転数以下に低下するまで待機する。所定高速回転数は、ニュートラルから指令された高速段の変速段に切換える油圧漸増制御を行った場合に成立する可能性がある変速段のギヤ比又は指令された変速段のギヤ比を乗じるとエンジンの制限回転数と等しくなる。これにより、指令された変速段よりギヤ比が大きい変速段が成立しても、エンジンが駆動輪によって過回転数で連れ回りされることがない。
【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、走行中に変速機構をニュートラルから高速段のいずれかの変速段に切換える指令が送出されたとき、出力軸の回転数が所定低速回転数より大きい場合に、高速段係合要素が早急に係合されると、低速段係合要素用の油圧サーボ装置に印加される油圧がカットオフ弁によって遮断されるので、低速段の成立を早急に解消することができ、例えば走行中にマニュアルバルブがニュートラルレンジＮにシフトされた後に、前進走行レンジDに戻されても、エンジンが駆動輪によって過回転数で連れ回りされることがない。
【００２０】
【実施の形態】
以下、図面に基づいて本発明の実施形態を説明する。図１は本発明に係る自動変速機の制御装置により制御される自動変速機１０の一例を示すスケルトン図で、自動変速機１０は、図略のエンジンが回転連結された流体伝動装置としてのトルクコンバータ１１及びトルクコンバータ１１から入力軸２０に入力された入力回転を変速して出力軸２１に出力する前進６速、後進１速の変速機構１２から構成されている。トルクコンバータ１１は、ポンプインペラ１３、タービンランナ１４、ステータ１５、ステータ１５を変速機構１２のケース１６に一方向の回転のみ許容して支承するワンウェイクラッチ１７、ワンウェイクラッチ１７のインナレースをケース１６に固定するステータシャフト１８を備えている。１９はポンプインペラ１３とタービンランナ１４とを直結するロックアップクラッチである。
【００２１】
変速機構１２の減速プラネタリギヤＧ１は、シングルピニオン型で、第１リングギヤＲ１が入力軸２０に連結され、第１サンギヤＳ１がケース１６に固定されて反力を受け、第１キャリヤＣ１に支承されたピニオンが第１リングギヤＲ１と第１サンギヤＳ１とに噛合されている。変速機構１２の主要部である変速プラネタリギヤＧは、ダブルピニオン型で、大径の第２サンギヤＳ２、小径の第３サンギヤＳ３、第２サンギヤＳ２に直接噛合するとともに第３サンギヤＳ３にピニオンＰ３を介して噛合するロングピニオンＰ２、ロングピニオンＰ２及びピニオンＰ３を支持する第２キャリヤＣ２（Ｃ３）及びロングピニオンＰ２と噛合し出力軸２１に連結された第２リングギヤＲ２（Ｒ３）から構成されている。
【００２２】
減速プラネタリギヤＧ１の第１キャリヤＣ１が第１クラッチC-1を介して変速プラネタリギヤＧの第３サンギヤＳ３に連結されるとともに、第３クラッチC-3を介して第２サンギヤＳ２に連結されている。変速プラネタリギヤＧの第２サンギヤＳ２は第１ブレーキB-1に連結され、第２キャリヤＣ２（Ｃ３）は第２クラッチC-2を介して入力軸２０に連結されるとともに、ケース１６に支持されたワンウェイクラッチF-1及び第２ブレーキB-2に並列に連結されている。
【００２３】
自動変速機１０の各クラッチ、ブレーキ及びワンウェイクラッチの係合、解放と各変速段との関係は図２の係合表に示すようになる。係合表における○印は係合、無印は解放、△印はエンジンブレーキ時のみの係合を示す。図３は各クラッチ、ブレーキ及びワンウェイクラッチの係合により成立される変速段と、そのときのプラネタリギヤＧ，Ｇ１の各要素の回転数比との関係を示す速度線図である。
【００２４】
図２，３から明らかなように、第１速段（１ｓｔ）は、第１クラッチC-1の係合とワンウェイクラッチF-1の自動係合によって達成される。入力軸２０の回転が減速プラネタリギヤＧ１によって減速された第１キャリヤＣ１の回転が、第１クラッチC-1により変速プラネタリギヤＧの第３サンギヤＳ３に入力され、ワンウェイクラッチF-1によって逆転を阻止された第２キャリヤＣ２（Ｃ３）が反力を受け、第２リングギヤＲ２（Ｒ３）が最大ギヤ比で減速回転されて出力軸２１に出力する。
【００２５】
第２速段（２ｎｄ）は、第１クラッチC-1と第１ブレーキB-1の係合によって達成される。入力軸２０の回転が減速プラネタリギヤＧ１によって減速された第１キャリヤＣ１の回転が、第１クラッチC-1経由で変速プラネタリギヤＧの第３サンギヤＳ３に入力され、第１ブレーキB-1の係合によって回転を阻止された第２サンギヤＳ２が反力を受け、第２リングギヤＲ２（Ｒ３）が第２速段に減速回転されて出力軸２１に出力する。このときのギヤ比は、図３に示すように、第１速段（１ｓｔ）より小さくなる。
【００２６】
第３速段（３ｒｄ）は、第１及び第３クラッチC-1，C-3の係合によって達成される。入力軸２０の回転が減速プラネタリギヤＧ１によって減速された第１キャリヤＣ１の回転が、第１及び第３クラッチC-1，C-3により第３及び第２サンギヤＳ３，Ｓ２に同時に入力されて変速プラネタリギヤＧが直結状態となり、第２リングギヤＲ２（Ｒ３）が第１キャリヤC1と同一回転数で回転されて出力軸２１に出力する。
【００２７】
第４速段（４ｔｈ）は、第１及び第２クラッチC-1，C-2の係合によって達成される。入力軸２０の回転が第２クラッチC-2により変速プラネタリギヤＧの第２キャリヤＣ２（Ｃ３）に直接入力され、入力軸２０の回転が減速プラネタリギヤＧ１によって減速された第１キャリヤＣ１の回転が、第１クラッチC-1により変速プラネタリギヤＧの第３サンギヤＳ３に入力され、第２リングギヤＲ２（Ｒ３）が入力軸２０と第１キャリヤＣ１との中間の回転数に減速されて出力軸２１に出力する。
【００２８】
第５速段（５ｔｈ）は、第２及び第３クラッチC-2，C-3の係合により達成される。入力軸２０の回転が第２クラッチC-2により変速プラネタリギヤＧの第２キャリヤＣ２（Ｃ３）に直接入力され、入力軸２０の回転が減速プラネタリギヤＧ１によって減速された第１キャリヤＣ１の回転が、第３クラッチC-3により変速プラネタリギヤＧの第２サンギヤＳ２に入力され、第２リングギヤＲ２（Ｒ３）が第５速段に増速回転されて出力軸２１に出力する。
【００２９】
第６速段（６ｔｈ）は、第２クラッチC-2と第１ブレーキB-1との係合により達成される。入力軸２０の回転が第２クラッチC-2により変速プラネタリギヤＧの第２キャリヤＣ２（Ｃ３）に直接入力され、第１ブレーキB-1の係合によって回転を阻止された第２サンギヤＳ２が反力を受け、第２リングギヤＲ２（Ｒ３）が第６速段に増速回転されて出力軸２１に出力する。
【００３０】
後進段（Ｒ）は、第３クラッチC-3と第２ブレーキB-2との係合によって達成される。入力軸２０の回転が減速プラネタリＧ１によって減速された第１キャリヤＣ１の回転が、第３クラッチC-3経由で変速プラネタリギヤＧの第２サンギヤＳ２に入力され、第２ブレーキB-2の係合によって回転を阻止された第２キャリヤＣ２（Ｃ３）が反力を受け、第２リングギヤＲ２（Ｒ３）が逆転されて出力軸２１に出力する。
【００３１】
上記自動変速機１０においては、第１乃至第３クラッチC-1〜C-3及び第１ブレーキB-1が係合要素である。ここで係合要素とは前進時の走行レンジＤ又は後進時の走行レンジＲが選択されたときに係合されるクラッチ又はブレーキを含む概念である。第１クラッチC-1が、変速機構１２の低速段である第１〜第３速段の成立時に常時係合される低速段係合要素であり、第２クラッチC-2が、高速段である第４〜第６速段の成立時に常時係合される高速段係合要素である。また、走行レンジが選択された状態でブレーキの制動力によって車両が第１速段状態で停止し、ニュートラル制御されて第１クラッチC-1が解放された状態になると第１速段から第２速段が選択されるため発進時にブレーキが離された場合、登り坂で車両に後退方向の力が作用しても第１ブレーキB-1が出力軸２１の逆転を阻止するので、車両は後退することがなく、第１クラッチC-1が係合し始めて駆動力を伝達すると第１速段に切換えられて車両は円滑に発進する。なお、エンジンブレーキが必要な時には第２ブレーキB-2が係合され、第２キャリヤＣ２（Ｃ３）が正回転を阻止されて、出力軸２１からの回転が第３サンギヤＳ３、第１クラッチC-1、減速プラネタリＧ１、トルクコンバータ１１を経由してエンジンに伝達され、エンジンブレーキがかかる。
【００３２】
次に、第１クラッチC-1の油圧駆動部に給排される出力油圧を送出する油圧サーボ装置２６を図４に基づいて説明する。２５は運転者がシフトレバーを操作してニュートラルＮ、前進走行レンジD、後進走行レンジＲに手動で切換えるマニュアルバルブで、ポートＰＬにオイルポンプから油圧が圧力制御弁４７によって所定圧に制御されたライン圧が供給されている。マニュアルバルブ２５が前進走行レンジＤにシフトされたときポートＰＬと連通されるポートＤには、第１クラッチC-1の油圧駆動部に供給される出力油圧を出力する油圧サーボ装置２６の増幅弁２７の入力ポート２８及び切換弁２９のライン圧ポート３０が夫々接続されている。３１はオイルポンプＰからのライン圧が減圧弁を介して供給されるソレノイドモジュレータバルブで、所定圧に制御した出力油圧を油圧サーボ装置２６のリニアソレノイド調圧弁３２の入力ポート３３及び切換弁２９のポート３４に供給する。
【００３３】
リニアソレノイド調圧弁３２は、ソレノイドSLC-1が後述する制御装置から供給される制御信号である制御電流に応じて作動して弁体３６を圧縮バネ３７のバネ力とバランスする位置まで移動し、入力ポート３３から流入する所定圧に制御された油圧を絞って制御装置からの制御電流の増大につれて減少する制御油圧を出力ポート３８に生成する。リニアソレノイド調圧弁３２の出力ポート３８は、増幅弁２７の制御ポート３９に接続されるとともに、切換弁２９の切換ポート４０に接続されている。増幅弁２７は、弁体４９が制御ポート３９から供給されて弁体４９の大径端面に作用するリニアソレノイド調圧弁３２の制御油圧による軸力が弁体４９の小径端面に作用する圧縮バネ４１のバネ力とフィードバック油圧による軸力とが釣り合う位置に移動され、入力ポート２８に供給されたライン圧を制御電流の増大につれて減少するリニアソレノイド調圧弁３２の制御油圧に応じた出力油圧Pcを出力し、出力ポート４２から切換弁２９の入力ポート４３に供給する。
【００３４】
切換弁２９は、弁体４５が図示右半分位置にシフトされると、入力ポート４３を出力ポート４４に連通し、増幅弁２７からの出力油圧Pcを第１クラッチC-1の油圧駆動部に供給し、弁体４５が図示左半分位置にシフトされると、ライン圧ポート３０を出力ポート４４に連通し、マニュアルバルブ２５のポートＤからのライン圧を第１クラッチC-1の油圧駆動部に供給し、第１クラッチC-1をライン圧によって係合状態に維持する。前進６段の各変速段を成立するために係脱される第１乃至第３クラッチC-1〜C-3及び第１ブレーキB-1の各油圧駆動部に油圧を給排する油圧サーボ装置２６は全て同様の構成であるので、図５に示すこれら係合要素を係脱する油圧回路においては、各係合要素に油圧を給排する油圧サーボ装置２６、増幅弁２７、切換弁２９及びリニアソレノイド調圧弁３２に同一参照番号を付して表示し、リニアソレノイドについては、SLの後に各係合要素を示す参照符号を付加してSLC-1，SLC-2，SLC-3，SLB-1と表示する。
【００３５】
図５に示すように、オイルポンプＰからの油圧が圧力制御弁４７によって所定圧に制御されたライン圧が供給されるマニュアルバルブ２５のポートDは、カットオフ弁５０を介して第１クラッチC-1用の油圧サーボ装置２６の増幅弁２７の入力ポート２８及び切換弁２９のライン圧ポート３０に並列に接続されている。カットオフ弁５０の制御ポートには、第2クラッチC-2に供給される油圧及び第３クラッチC-3又は第１ブレーキB-1に供給される油圧が供給され、第２クラッチC-2に供給される油圧が高く、且つ第３クラッチC-3又は第１ブレーキB-1に供給される油圧が高いとき、ポートDから第１クラッチC-1用の油圧サーボ装置２６への油圧の供給を遮断する。マニュアルバルブ２５のポートDは、カットオフ弁５１を介して第２クラッチC-2用の油圧サーボ装置２６の増幅弁２７の入力ポート２８及び切換弁２９のライン圧ポート３０に並列に接続されている。カットオフ弁５１の制御ポートには、第2クラッチC-2に供給される油圧及びソレノイド弁４８の出力油圧が供給され、第２クラッチC-2に供給される油圧が高いとき、又はソレノイド弁４８の出力油圧が高いとき、ポートDから第２クラッチC-2用の油圧サーボ装置２６への油圧の供給を許容する。
【００３６】
第３クラッチC-3及び第１ブレーキB-1用のサーボ装置２６の切換弁２９は、リニアソレノイド調圧弁３２のソレノイドSLC-3又はSLB-1が付勢されて出力ポート３８から切換ポート４０に供給される圧力が低くなると連通される弁５２，５３が付加されている。圧力制御弁４７からライン圧を供給されるカットオフ弁５４は弁５２，５３に並列に接続され、弁５３は第３クラッチC-3用の油圧サーボ装置２６の増幅弁２７の入力ポート２８及び切換弁２９のライン圧ポート３０に並列に接続され、弁５２は第１ブレーキB-1用の油圧サーボ装置２６の増幅弁２７の入力ポート２８及び切換弁２９のライン圧ポート３０に並列に接続されている。５５はソレノイドSLC-3を有するリニアソレノイド調圧弁３２の出力ポート３８と第３クラッチC-3用の増幅弁２７の制御ポート３９及び切換弁２９の切換ポート４０との間に接続されたカットオフ弁で、カットオフ弁５５の制御ポートには、第2クラッチC-2に供給される油圧及び第１ブレーキB-1に供給される油圧が供給され、第２クラッチC-2及び第１ブレーキB-1に供給される油圧が高いとき、出力ポート３８から制御ポート３９及び切換ポート４０への油圧の供給を遮断する。５６はソレノイドSLB-1を有するリニアソレノイド調圧弁３２の出力ポート３８と第１ブレーキB-1用の増幅弁２７の制御ポート３９及び切換弁２９の切換ポート４０との間に接続されたカットオフ弁で、カットオフ弁５６の制御ポートには、カットオフ弁５５の出力側が接続され、カットオフ弁５５の出力側の油圧が高いとき、出力ポート３８から制御ポート３９及び切換ポート４０への油圧の供給を遮断する。なお、５７は第３クラッチC-3及び第１ブレーキB-1の油圧を、お互いを連通することなくカットオフ弁５０の制御ポートに伝達するバルブである。
【００３７】
自動変速機の制御装置を図６に示すブロック図に基づいて説明する。CPUを内蔵した制御装置６０は、エンジンの回転が伝達されるトルクコンバータ１１のエンジン側回転数Neを検出するエンジン側回転数センサ６１、入力軸２０の回転数Niを検出する入力軸回転数センサ６２、出力軸２１の回転数Ｎｖを検出する出力軸回転数センサ６３、マニュアルバルブ２５が前進走行レンジＤ、ニュートラルレンジＮにシフトされているとき、検出信号Dr，Nrを送出するレンジ位置センサ６４、アクセルの踏み込み量Ssを検出するスロットル開度センサ６５等から各検出信号が入力され、これら検出信号に基づいて制御信号である制御電流を各油圧サーボ装置２６のリニアソレノイド調圧弁３２に出力するとともに、低速段成立時に常時係合される第１クラッチC-1が異常係合しているときに、ニュートラルレンジＮから前進走行レンジＤにおける高速段の中のいずれかの変速段に切換える指令が送出されたとき、エンジンが駆動輪によって過回転数で連れ回りされることを防止する図７に示す制御プログラム７０が記憶されている。
【００３８】
次に、本発明に係る自動変速機の制御装置の実施形態の作動を説明する。第５速段で走行中にマニュアルバルブ２５がニュートラルレンジＮにシフトされると、ソレノイドSLC-1〜SLC-3及びSLB-1はオン状態にされ、ポートＤから第1、第2クラッチC-1，C-2用の油圧サーボ装置２６への油圧の供給が遮断され、第１、第２クラッチC-1，C-2が解放される。ソレノイドSLC-3，SLB-1のオンによって第３クラッチC-3、第１ブレーキB-1用の油圧サーボ装置２６からの出力油圧Pcがゼロになり第３クラッチC-1、第１ブレーキB-1が解放されてニュートラルとなる。
【００３９】
その後に前進走行レンジＤに戻され、ニュートラルレンジＮから前進走行レンジＤにおける高速段のいずれかの変速段、例えば第５速段に切換える指令が送出されると、制御装置６０は正常時においては、油圧サーボ装置２６から第２クラッチC-2に供給される油圧を漸増する漸増制御であるND制御を行ない、ソレノイドSLC-3をオフ状態にして第３クラッチC-3を係合し、第２クラッチC-2を係合ショックが発生しないように係合して第５速段を成立する。ND制御においてソレノイド弁４８が所定時間オン状態にされると、制御ポートに油圧が供給されて第２カットオフ弁５１がシフトされ、油圧がポートＤから第２クラッチC-2用の油圧サーボ装置２６に供給される。そして、図８に示すように、制御装置６０は、第２クラッチC-2のクラッチ板間の隙間を早くつめるために、第２クラッチC-2用の油圧サーボ装置２６のソレノイドSLC-2に供給する制御電流Ｉをパルス状に低下して第２クラッチC-2の油圧駆動部にパルス状の初期油圧を供給した後に、制御電流を若干増加して第２クラッチC-2がクラッチ板の接触によりスリップを開始する位置まで油圧駆動部のピストンを移動させるために必要なストローク圧を供給し、次に制御電流Ｉを漸減して出力油圧Pcを漸増し、第２クラッチC-2をショックが発生しないように係合する。その後に制御電流は０に低下され、第２クラッチC-2はライン圧によって係合状態に維持される。
【００４０】
このときに、第１クラッチC-1用の油圧サーボ装置２６が、ソレノイドSLC-1をオン状態にしても出力油圧を送出してクラッチC-1を異常係合する故障の場合、マニュアルバルブ２５が前進走行レンジＤにシフトされると、ポートＤから第１クラッチC-1用の油圧サーボ装置２６に油圧が供給され、第１クラッチC-1が異常係合され、第２クラッチC-2が係合されるまでの間に第３速段が成立する。第２クラッチC-2用の油圧サーボ装置２６からの出力油圧Pcが上昇して第２クラッチC-2が係合されると、第５速段が成立する。第２、第３クラッチC-2，C-3に供給される油圧が制御ポートに供給されて第１カットオフ弁５０がシフトされ、第１クラッチC-1用の油圧サーボ装置２６がポートDから遮断され、第１クラッチC-1が解放される。このように、カットオフ弁５０は、第２クラッチC-2用の油圧サーボ装置２６に油圧が印加され第２クラッチC-2が係合されると、第１クラッチC-1用の油圧サーボ装置２６に印加される油圧を遮断する。
【００４１】
ND制御において、ソレノイドSLC-3をオフ状態にしたとき、第３クラッチC-3用の油圧サーボ装置２６の出力油圧の上昇が、第１、第２クラッチC-1，C-2用の油圧サーボ装置２６の出力油圧の上昇より何らかの理由で遅くなった場合、第１、第２クラッチC-1，C-2に供給される油圧が制御ポートに供給されてカットオフ弁５４がシフトされ、第３クラッチC-3、第１ブレーキB-1用の油圧サーボ装置２６へのライン圧の供給を遮断するので、第３クラッチC-3は係合されない。制御ポートに第３クラッチC-3から油圧が供給されないので、第１カットオフ弁５０が第１クラッチC-1用の油圧サーボ装置２６をポートDから遮断することがなく、第１クラッチC-1が係合されて第４速段が成立することもある。
【００４２】
制御装置６０は、第５速段で走行中にマニュアルバルブ２５がニュートラルレンジＮにシフトされた後、前進走行レンジＤに戻されたことをレンジ位置センサ６４からの検出信号Nr，Drによって認識すると、変速機構１２をニュートラルＮから第５速段に切換える指令を送出するとともに、変速機構１２が低速段にシフトダウンしてエンジンが過回転数で回転されることを防止するために、制御プログラム７０を実行する。ステップ７１で出力軸２１の回転数Nvが所定高速回転数RHより大きいか否か判定する。所定高速回転数RHは、図９に示す横軸に出力軸２１の回転数Nｖ、縦軸に入力軸２０の回転数Niを取ったNv‐Ni平面において、エンジンの制限回転数に対応する入力軸２０の回転数Rev limitを示す直線と、第４速段を示す直線４thとの交点の出力軸２１の回転数である。第４変速段は、ニュートラルＮから指令された第５速段に切換えるND制御を行った場合に成立する可能性がある変速段である。回転数Nvが所定高速回転数RHより大きくて回転数Nv，NiがNv‐Ni平面の領域▲４▼に入っている場合、ステップ７２で回転数Nvが所定高速回転数RHより小さくなるまで待機する。これにより、ND制御で第４速段が成立してもエンジンが駆動輪によって制限回転数以上に連れ回りされることがない。
【００４３】
回転数Nvが所定高速回転数RHより小さい場合、ステップ７３で出力軸２１の回転数Nvが所定低速回転数RLより小さいか否か判定する。回転数Nvが所定低速回転数RLより小さくてNv‐Ni平面の領域▲３▼に入っている場合、第３、第４速段が成立してもエンジンが駆動輪によって制限回転数以上に連れ回りされることがないので、ステップ７４でND制御が通常通り行われる。所定低速回転数RLは、Nv‐Ni平面において、エンジンの制限回転数に対応する入力軸２０の回転数Rev limitを示す直線と、第３速段を示す直線３rdとの交点の出力軸２１の回転数である。
【００４４】
回転数Nvが所定低速回転数RLより大きい場合、ステップ７５で入力軸回転数センサ６２によって検出された入力軸２０の回転数Niが、出力軸回転数センサ６３によって検出された出力軸２１の回転数Nvに指令された変速段である第５速段のギヤ比Gr5を乗じた理想回転数（Nv×Gr5）より小さいか否か判定する。小さくてNv‐Ni平面の領域▲２▼に入っている場合、図８に示すように、ステップ７６で第５速段を指令してND制御が行なわれ、第２クラッチC-2用の油圧サーボ装置２６の出力油圧Pcが漸増して第２クラッチC-2が徐々に係合される。ステップ７７で入力軸２０の回転数Niが、出力軸２１の回転数Nvに第５速段のギヤ比Gr5を乗じた入力軸２０の理想回転数（Nv×Gr5）とほぼ等しくなるまで増加したことが判定されると、ステップ７８でND制御が中断され、ソレノイドSLC-2に印加される制御電流が０に低下され、第２クラッチC-2がライン圧によって係合されて第５速段が成立する。
【００４５】
入力軸２０の回転数Niが、理想回転数（Nv×Gr5）より大きくてNv‐Ni平面の領域▲１▼に入っている場合、第５速段を指令したND制御中に第３速段が成立すると、出力軸２１の回転数Nvに第３速段のギヤ比Gr3を乗じた予想回転数（Nv×Gr3）がエンジンの制限回転数に対応する入力軸２０の回転数Rev limitを超え、エンジンが駆動輪によって過回転数で回転される可能性があるので、ステップ７９でエンジンのトルクリミテーション制御を行ってエンジンの出力トルクを制限し、ステップ８０で入力軸２０の回転数Niが出力軸２１の回転数Nvに第５速段のギヤ比Gr5を乗じた入力軸２０の理想回転数（Nv×Gr5）とほぼ等しくなるまで低下したことが判定されると、ステップ８１で、ND制御を行うことなく、図１０に示すように、第２クラッチC-2に油圧を給排する油圧サーボ装置２６の出力油圧を急峻に高くして第２クラッチC-2を早急に係合する油圧急峻制御を行う。
【００４６】
第６速段で走行中にマニュアルバルブ２５がニュートラルレンジにシフトされた後に、前進走行レンジＤに戻された場合、第６速段を指令したND制御でソレノイドSLB-1がオフ状態にされて第１ブレーキB-1が係合され、ソレノイドSLC-2用の油圧サーボ装置２６の出力油圧Pcが漸増されて第２クラッチC-2が係合するまでの間に第２速段が成立する以外は、上述の第５速段の場合と同様であるので、詳細な説明は省略する。
【００４７】
上記実施形態においては、エンジンの制限回転数に対応する入力軸２０の回転数Rev limitを示す直線と、ニュートラルＮから指令された第５速段に切換えるND制御を行った場合に成立する可能性がある第４速段を示す直線４thとの交点の出力軸２１の回転数を所定高速回転数RHとしているが、指令した変速段と異なる変速段が成立する可能性が無い場合、又は成立する可能性がある変速段の方がギヤ比が小さい場合は、回転数Rev limitを示す直線と、指令した変速段を示す直線との交点の出力軸２１の回転数を所定高速回転数RHとする。
【００４８】
また、上記実施形態においては、入力軸２０の回転数Niが、出力軸２１の回転数Nvに指令された第５速段のギヤ比Gr5を乗じた理想回転数より大きいか否か判定しているが、入力軸回転数センサ６２、出力軸回転数センサ６３によって検出された入力軸２０の回転数Niと出力軸２１の回転数Nvとの比を実際のギヤ比として演算し、この実際のギヤ比が指令された第５速段のギヤ比Gr5より大きいか否か判定するようにしてもよい。
【図面の簡単な説明】
【図１】本発明に係る自動変速機の制御装置により制御される自動変速機の実施形態のスケルトン図。
【図２】自動変速機の各変速段におけるクラッチ、ブレーキの係合表。
【図３】自動変速機の各変速段におけるプラネタリギヤの各要素の回転数比を示す速度線図。
【図４】第１クラッチC-1を係脱する油圧サーボ装置を示す図。
【図５】油圧サーボ装置に油圧を給排するための油圧回路図。
【図６】自動変速機の制御装置を示すブロック図。
【図７】制御プログラムのフロー図。
【図８】ニュートラルレンジから前進走行レンジにシフトしたときに、ND制御を行った場合の各クラッチの係合状態を示す図。
【図９】出力軸及び入力軸の回転数に適した制御を行うために両回転数の領域を区分した図。
【図１０】ニュートラルレンジから前進走行レンジにシフトしたときに、ND制御を行わない場合の各クラッチの係合状態を示す図。
【符号の説明】
１０・・・自動変速機、１１・・・トルクコンバータ（流体伝動装置）、１２・・・変速機構、１９・・・ロックアップクラッチ、２０・・・入力軸、２１・・・出力軸、２５・・・マニュアルバルブ、２６・・・油圧サーボ装置、２７・・・増幅弁、２９・・・切換弁、３２・・・リニアソレノイド調圧弁、SLC-1〜SLC-3，SLB-1，SLB-2・・・リニアソレノイド、３６，４９・・・弁体、４７・・・圧力制御弁、４８・・・ソレノイド弁、５１，５４〜５６・・・カットオフ弁、５２，５３・・・弁、６０・・・制御装置、６１・・・エンジン側回転数センサ、６２・・・入力軸回転数センサ、６３・・・出力軸回転数センサ、６４・・・レンジ位置センサ、６５・・・スロットル開度センサ、C-1・・・第１クラッチ、C-2・・・第２クラッチ、C-3・・・第３クラッチ、B-1・・・第１ブレーキ、B-2・・・第２ブレーキ、７０・・・制御プログラム。[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a control device for an automatic transmission that prevents an engine from rotating at an overspeed when a manual valve of an automatic transmission is shifted to a neutral range during traveling and returned to a forward traveling range. is there.
[0002]
[Prior art]
When the vehicle is traveling, the shift lever is operated to shift the manual valve from the forward travel range to the neutral range and return to the forward travel range. Conventionally, in hydraulic gradual increase control performed when the manual valve is switched from the neutral range to the forward travel range, the engagement shock is reduced when the second clutch is engaged to establish the fifth or sixth speed. Therefore, the output hydraulic pressure of the hydraulic servo device supplied to the hydraulic drive unit of the second clutch is gradually increased.
[0003]
[Problems to be solved by the invention]
However, when a hydraulic servo device that supplies and discharges hydraulic pressure to and from the first clutch that is always engaged when the low speed stage is established fails even when the linear solenoid is turned on and the output hydraulic pressure is output and the first clutch is abnormally engaged. For example, when the manual valve is shifted to the neutral range N during traveling at the fifth speed, the hydraulic servo device for the first and second clutches C-1 and C-2 from the port D of the manual valve 25 in FIG. 26 is cut off, the first and second clutches C-1 and C-2 are released, the solenoids SLC-3 and SLB-1 are turned on, and the third clutch C-3 and the first clutch The hydraulic pressure output from the hydraulic servo device 26 for the brake B-1 becomes zero, and the third clutch C-1 and the first brake B-1 are released to become neutral. The cut-off valve 51 that has stopped supplying hydraulic pressure to the control port shuts off the hydraulic servo device 26 for the second clutch C-2 from the port D. After that, when the manual valve 25 is shifted to the forward travel range D and the fifth speed is commanded to perform hydraulic pressure gradual increase control, the solenoid SLC-3 is turned off and the third clutch C-3 is engaged. The hydraulic pressure is supplied from the port D to the hydraulic servo device 26 for the first clutch C-1, the first clutch C-1 is abnormally engaged, and the third speed is established. The solenoid valve 48 is turned on for a predetermined time, the cut-off valve 51 is shifted, the hydraulic pressure is supplied from the port D to the hydraulic servo device 26 for the second clutch C-2, and the control current applied to the solenoid SLC-2 Is gradually reduced to an off state, and the second clutch C-2 is engaged with a slight delay. The hydraulic pressure supplied to the second and third clutches C-2 and C-3 is supplied to the control port, the cut-off valve 50 is shifted, and the hydraulic servo device 26 for the first clutch C-1 is disconnected from the port D. Then, the first clutch C-1 is released.
[0004]
In this way, if the fifth clutch is commanded and hydraulic pressure gradual increase control is performed in the event of a failure in which the first clutch C-1 is abnormally engaged, the third gear is established during the control, and the engine is driven. There was a possibility that it would be rotated by the wheel and rotated at excessive speed. Similarly, when the sixth speed is commanded and hydraulic pressure gradual increase control is performed, the second speed is established in the middle of the control.
[0005]
The present invention has been made to solve such a conventional problem, and even when the failure is such that the engagement element that is always engaged when the low speed stage of the automatic transmission is established is abnormally engaged, the hydraulic pressure increase control is performed. It is to prevent the engine from being driven by the drive wheel at an excessive speed.
[0006]
[Means for Solving the Problems]
In order to solve the above-described problem, the structural feature of the invention according to claim 1 is that a transmission mechanism that shifts the rotation of the input shaft and outputs it to the output shaft, and each gear stage of the transmission mechanism are established. A plurality of engaging elements to be engaged / disengaged, a plurality of hydraulic servo devices for supplying / discharging output hydraulic pressure corresponding to an electric signal supplied to the solenoid to disengage the engaging elements, and an input shaft of the transmission mechanism In the control device for an automatic transmission provided with an input shaft rotation speed detection device and an output shaft rotation speed detection device for detecting the rotation speed of the transmission shaft and the rotation speed of the output shaft, respectively, A high-speed stage engaging element that is always engaged when a stage is established, a low-speed stage engaging element that is always engaged when a low-stage stage is established, and one that is engaged with the high-speed stage engaging element or the low-speed stage engaging element. third engagement to establish the respective speeds of the high-speed stage or the low speed stage Te The low speed stage engaging element includes an element and a fourth engaging element, and the high speed stage engaging element and the third engaging element or the fourth engaging element are engaged to establish the high speed stage. The control device has a shut-off means for shutting off the supply of hydraulic pressure to the engine, and the controller rotates the output shaft when a command is issued to switch the speed change mechanism from neutral to any of the high speed stages during travel. Means for determining whether or not the number is smaller than a predetermined low speed, and if so, the hydraulic pressure supplied from the hydraulic servo device to the high-speed stage engaging element is gradually increased, and the speed change mechanism is instructed from neutral. Means for performing hydraulic pressure gradual increase control for switching to the gear position; and, if large, the output hydraulic pressure of the hydraulic servo device that supplies and discharges hydraulic pressure to and from the high speed gear engaging element without interrupting or performing the hydraulic pressure gradual increase control is sharply increased. The high speed stage Is that comprising means for performing hydraulic steep control to quickly engage the coupling elements.
[0007]
According to a second aspect of the present invention, in the control device for an automatic transmission according to the first aspect, the predetermined low-speed rotation speed is obtained when the low-speed stage engagement element is abnormally engaged. When the gear ratio of the low speed stage established in the middle of the hydraulic pressure increase control is multiplied, the rotation speed of the output shaft becomes equal to the engine speed limit.
[0008]
According to a third aspect of the present invention, in the control device for an automatic transmission according to the first or second aspect, the control device may be configured such that the rotational speed of the output shaft is greater than the predetermined low-speed rotational speed. Whether the rotational speed of the input shaft is larger than the ideal rotational speed obtained by multiplying the rotational speed of the output shaft by the gear ratio of the commanded gear, or the rotational speed of the input shaft and the rotational speed of the output shaft, Means for determining whether or not the actual gear ratio is larger than the gear ratio of the commanded gear, and if so, engine torque limit control is performed so that the rotational speed of the input shaft is equal to the ideal rotational speed. When almost equal, the output hydraulic pressure of a hydraulic servo device that supplies and discharges hydraulic pressure to and from the high-speed stage engaging element is rapidly increased without performing the hydraulic pressure gradual increase control, and the high-speed stage engaging element is quickly engaged. It has a means to perform hydraulic steep control. .
[0009]
According to a fourth aspect of the present invention, in the control device for an automatic transmission according to the third aspect, the control device is configured so that the rotation speed of the output shaft is greater than the predetermined low-speed rotation speed. When the rotational speed of the input shaft or the actual gear ratio is smaller than the ideal rotational speed or the gear ratio of the commanded gear, the hydraulic pressure gradually increases when the rotational speed of the input shaft becomes substantially equal to the ideal rotational speed. Means for steeply increasing the output hydraulic pressure of a hydraulic servo device that interrupts control and supplies / discharges hydraulic pressure to / from the high speed stage engaging element to perform hydraulic steep control to quickly engage the high speed stage engaging element That is.
[0010]
The structural feature of the invention according to claim 5 is the control device for an automatic transmission according to any one of claims 1 to 4, wherein the control device has a rotational speed of the output shaft larger than a predetermined high speed rotational speed. And a means for waiting until the rotational speed of the output shaft drops below the predetermined high speed rotational speed.
[0011]
According to a sixth aspect of the present invention, there is provided a control device for an automatic transmission according to the fifth aspect, wherein the predetermined high-speed rotation speed is controlled by a gradual hydraulic pressure increase control for switching to a high-speed gear stage commanded from neutral. If it is performed, the output shaft speed becomes equal to the engine speed limit when multiplied by the gear ratio of the gear stage that may be established or the gear ratio of the commanded gear stage.
[0012]
The structural feature of the invention according to claim 7 is the control device for an automatic transmission according to any one of claims 1 to 6, wherein hydraulic pressure is applied to the hydraulic servo device for the high-speed stage engaging element, and the high-speed When the step engagement element is engaged, a cut-off valve that cuts off the hydraulic pressure applied to the hydraulic servo device for the low speed step engagement element is provided.
[0013]
[Operation and effect of the invention]
In the invention according to claim 1 configured as described above, when a command for switching the speed change mechanism from the neutral position to any one of the high speed stages is sent during traveling, the engine is driven by the drive wheels at an overspeed. In order to prevent rotation, it is determined whether or not the rotation speed of the output shaft is smaller than a predetermined low-speed rotation speed. If it is smaller, the hydraulic pressure supplied from the hydraulic servo device to the high-speed engagement element is gradually increased. Hydraulic gradual increase control is performed to switch the speed change mechanism to the gear position commanded from neutral. If larger, the hydraulic steep control for quickly engaging the high-speed engaging element is performed without interrupting or performing the hydraulic gradual increase control. As a result, for example, even if the manual valve is shifted to the neutral range N during traveling and then returned to the forward traveling range D, the engine is not rotated by the drive wheels at the excessive speed.
[0014]
In the invention according to claim 2 configured as described above, when the low-speed stage engaging element is abnormally engaged, if the low-speed gear ratio established in the middle of the hydraulic pressure increase control is multiplied, the engine speed limit When the rotation speed of the output shaft is smaller than the predetermined low-speed rotation speed that is equal to the number, the hydraulic pressure gradually increasing control is performed. When the rotation speed is large, the high-speed engagement element is quickly engaged. It is possible to surely prevent being accompanied by excessive rotation speed.
[0015]
In the invention according to claim 3 configured as described above, when the rotational speed of the input shaft is larger than the ideal rotational speed obtained by multiplying the rotational speed of the output shaft by the gear ratio of the commanded gear, in other words, input If the actual gear ratio between the rotational speed of the shaft and the rotational speed of the output shaft is greater than the gear ratio of the commanded gear, the engine torque limitation control is performed and the rotational speed of the input shaft is almost equal to the ideal rotational speed. When they are equal, hydraulic steep control is performed in which the output hydraulic pressure of the hydraulic servo device that supplies and discharges hydraulic pressure to and from the high-speed engaging element is sharply increased and the high-speed engaging element is quickly engaged without performing hydraulic gradual increase control. Therefore, even if the low speed stage is established during the hydraulic pressure gradual increase control, the engine is not rotated by the drive wheel at the excessive rotation speed.
[0016]
In the invention according to claim 4 configured as described above, when the rotational speed of the input shaft is smaller than the ideal rotational speed obtained by multiplying the rotational speed of the output shaft by the commanded gear ratio, the rotational speed of the input shaft. When the engine speed is almost equal to the ideal rotation speed, the hydraulic pressure increase control is interrupted and the output hydraulic pressure of the hydraulic servo device that supplies and discharges the hydraulic pressure to and from the high-speed engagement element is sharply increased to quickly engage the high-speed engagement element. Since the combined hydraulic steep control is performed, even if the low speed stage is established in the middle of the gradual increase control, the engine is not driven by the drive wheel at the excessive rotation speed.
[0017]
In the invention according to claim 5 configured as described above, when the rotational speed of the output shaft is larger than the predetermined high-speed rotational speed, the process waits until the rotational speed of the output shaft decreases below the predetermined high-speed rotational speed. It is possible to reliably prevent the engine from being driven by the drive wheel at an excessive speed.
[0018]
In the invention according to claim 6 configured as described above, when the rotation speed of the output shaft is larger than the predetermined high-speed rotation speed, the process waits until the rotation speed decreases below the predetermined high-speed rotation speed. When the predetermined high-speed rotation speed is multiplied by the gear ratio of the shift speed or the gear ratio of the commanded shift stage, which is likely to be established when hydraulic gradual increase control for switching to the high-speed shift stage commanded from neutral is performed, the engine It becomes equal to the speed limit. As a result, even when a gear stage having a larger gear ratio than the commanded gear stage is established, the engine is not driven by the drive wheels at an excessive speed.
[0019]
In the invention according to claim 7 configured as described above, when a command for switching the speed change mechanism from neutral to any one of the high speed stages is sent during traveling, the rotation speed of the output shaft is the predetermined low speed rotation speed. If it is larger, if the high-speed stage engaging element is quickly engaged, the hydraulic pressure applied to the hydraulic servo device for the low-speed stage engaging element is cut off by the cutoff valve. For example, even if the manual valve is shifted to the neutral range N during traveling and then returned to the forward traveling range D, the engine is not driven by the drive wheels at an excessive speed.
[0020]
[Embodiment]
Hereinafter, embodiments of the present invention will be described with reference to the drawings. FIG. 1 is a skeleton diagram showing an example of an automatic transmission 10 controlled by a control device for an automatic transmission according to the present invention. The automatic transmission 10 is a torque as a fluid transmission device in which an unillustrated engine is rotationally connected. The transmission 11 includes a forward 6-speed and a reverse 1-speed transmission mechanism 12 that shifts the input rotation input from the converter 11 and the torque converter 11 to the input shaft 20 and outputs the input rotation to the output shaft 21. The torque converter 11 includes a pump impeller 13, a turbine runner 14, a stator 15, and a stator 15, a one-way clutch 17 that supports and supports the case 16 of the speed change mechanism 12 in only one direction, and an inner race of the one-way clutch 17 in the case 16. A stator shaft 18 to be fixed is provided. A lockup clutch 19 directly connects the pump impeller 13 and the turbine runner 14.
[0021]
The speed reduction planetary gear G1 of the speed change mechanism 12 is a single pinion type, the first ring gear R1 is connected to the input shaft 20, the first sun gear S1 is fixed to the case 16 and receives a reaction force, and is supported by the first carrier C1. The pinion is meshed with the first ring gear R1 and the first sun gear S1. A speed change planetary gear G that is a main part of the speed change mechanism 12 is of a double pinion type and directly meshes with the second sun gear S2 having a large diameter, the third sun gear S3 having a small diameter, and the second sun gear S2, and the pinion P3 is fitted to the third sun gear S3. And a second ring gear R2 (R3) engaged with the output shaft 21 and engaged with the long pinion P2, the second carrier C2 (C3) supporting the long pinion P2 and the pinion P3, and the long pinion P2. .
[0022]
The first carrier C1 of the reduction planetary gear G1 is connected to the third sun gear S3 of the transmission planetary gear G via the first clutch C-1, and is connected to the second sun gear S2 via the third clutch C-3. . The second sun gear S2 of the transmission planetary gear G is connected to the first brake B-1, and the second carrier C2 (C3) is connected to the input shaft 20 via the second clutch C-2 and supported by the case 16. The one-way clutch F-1 and the second brake B-2 are connected in parallel.
[0023]
The relationship between the engagement and release of each clutch, brake and one-way clutch of the automatic transmission 10 and each gear stage is as shown in the engagement table of FIG. In the engagement table, ◯ indicates engagement, no mark indicates release, and Δ indicates engagement only during engine braking. FIG. 3 is a velocity diagram showing the relationship between the gear stage established by engagement of each clutch, brake and one-way clutch, and the rotational speed ratio of each element of the planetary gears G and G1 at that time.
[0024]
As is apparent from FIGS. 2 and 3, the first speed (1st) is achieved by engagement of the first clutch C-1 and automatic engagement of the one-way clutch F-1. The rotation of the first carrier C1 whose rotation of the input shaft 20 is decelerated by the reduction planetary gear G1 is input to the third sun gear S3 of the transmission planetary gear G by the first clutch C-1, and the reverse rotation is prevented by the one-way clutch F-1. The second carrier C2 (C3) receives the reaction force, and the second ring gear R2 (R3) is decelerated and rotated at the maximum gear ratio and is output to the output shaft 21.
[0025]
The second speed (2nd) is achieved by engagement of the first clutch C-1 and the first brake B-1. The rotation of the first carrier C1 whose rotation of the input shaft 20 is decelerated by the reduction planetary gear G1 is input to the third sun gear S3 of the transmission planetary gear G via the first clutch C-1, and the first brake B-1 is engaged. The second sun gear S 2 whose rotation has been blocked by the force is subjected to a reaction force, and the second ring gear R 2 (R 3) is decelerated and rotated to the second speed stage and is output to the output shaft 21. The gear ratio at this time is smaller than the first gear (1st) as shown in FIG.
[0026]
The third speed (3rd) is achieved by engagement of the first and third clutches C-1 and C-3. The rotation of the first carrier C1 whose rotation of the input shaft 20 is decelerated by the speed reduction planetary gear G1 is simultaneously input to the third and second sun gears S3 and S2 by the first and third clutches C-1 and C-3 to change the speed. The planetary gear G is in a directly connected state, and the second ring gear R2 (R3) is rotated at the same rotational speed as the first carrier C1 and is output to the output shaft 21.
[0027]
The fourth speed (4th) is achieved by engagement of the first and second clutches C-1 and C-2. The rotation of the input shaft 20 is directly input to the second carrier C2 (C3) of the transmission planetary gear G by the second clutch C-2, and the rotation of the first carrier C1 whose rotation of the input shaft 20 is decelerated by the reduction planetary gear G1 is The first clutch C-1 is input to the third sun gear S3 of the transmission planetary gear G, and the second ring gear R2 (R3) is decelerated to an intermediate rotational speed between the input shaft 20 and the first carrier C1 and output to the output shaft 21. To do.
[0028]
The fifth speed (5th) is achieved by engagement of the second and third clutches C-2 and C-3. The rotation of the input shaft 20 is directly input to the second carrier C2 (C3) of the transmission planetary gear G by the second clutch C-2, and the rotation of the first carrier C1 whose rotation of the input shaft 20 is decelerated by the reduction planetary gear G1 is The third clutch C-3 is input to the second sun gear S2 of the transmission planetary gear G, and the second ring gear R2 (R3) is rotated to the fifth speed and output to the output shaft 21.
[0029]
The sixth speed (6th) is achieved by engagement of the second clutch C-2 and the first brake B-1. The rotation of the input shaft 20 is directly input to the second carrier C2 (C3) of the transmission planetary gear G by the second clutch C-2, and the second sun gear S2 whose rotation is blocked by the engagement of the first brake B-1 is counteracted. In response to the force, the second ring gear R2 (R3) is rotated to the sixth speed and output to the output shaft 21.
[0030]
The reverse speed (R) is achieved by engagement of the third clutch C-3 and the second brake B-2. The rotation of the first carrier C1 in which the rotation of the input shaft 20 is decelerated by the deceleration planetary G1 is input to the second sun gear S2 of the transmission planetary gear G via the third clutch C-3, and the second brake B-2 is engaged. The second carrier C 2 (C 3) that has been prevented from rotating by the force receives a reaction force, and the second ring gear R 2 (R 3) is reversely rotated and output to the output shaft 21.
[0031]
In the automatic transmission 10, the first to third clutches C-1 to C-3 and the first brake B-1 are engaging elements. Here, the engagement element is a concept including a clutch or a brake that is engaged when the travel range D for forward travel or the travel range R for reverse travel is selected. The first clutch C-1 is a low speed engagement element that is always engaged when the first to third speeds, which are the low speed stages of the transmission mechanism 12, are established, and the second clutch C-2 is a high speed stage. This is a high-speed stage engaging element that is always engaged when a certain fourth to sixth speed stage is established. In addition, when the travel range is selected, the vehicle is stopped in the first speed state by the braking force of the brake, and when the neutral control is performed and the first clutch C-1 is released, the second speed is changed from the first speed stage to the second speed stage. If the brake is released at the start because the speed is selected, the first brake B-1 prevents the output shaft 21 from reversing even if a reverse force is applied to the vehicle on the uphill, so the vehicle moves backward. When the first clutch C-1 begins to be engaged and the driving force is transmitted, the vehicle is switched to the first speed and the vehicle starts smoothly. When engine braking is required, the second brake B-2 is engaged, the second carrier C2 (C3) is prevented from rotating forward, and the rotation from the output shaft 21 is the third sun gear S3, the first clutch C. -1 is transmitted to the engine via the deceleration planetary G1 and the torque converter 11, and the engine brake is applied.
[0032]
Next, the hydraulic servo device 26 that sends out the output hydraulic pressure supplied to and discharged from the hydraulic drive unit of the first clutch C-1 will be described with reference to FIG. 25 is a manual valve that is manually switched to neutral N, forward travel range D, and reverse travel range R by the driver operating the shift lever. The hydraulic pressure from the oil pump to the port PL is controlled to a predetermined pressure by the pressure control valve 47. Line pressure is supplied. When the manual valve 25 is shifted to the forward travel range D, the port D communicated with the port PL has an amplifying valve of the hydraulic servo device 26 that outputs the output hydraulic pressure supplied to the hydraulic drive unit of the first clutch C-1. 27 input ports 28 and a line pressure port 30 of the switching valve 29 are connected to each other. 31 is a solenoid modulator valve to which the line pressure from the oil pump P is supplied via a pressure reducing valve. The output hydraulic pressure controlled to a predetermined pressure is applied to the input port 33 and the switching valve 29 of the linear solenoid pressure regulating valve 32 of the hydraulic servo device 26. Supply to port 34.
[0033]
The linear solenoid pressure regulating valve 32 operates according to a control current which is a control signal supplied from a control device described later by the solenoid SLC-1 and moves the valve body 36 to a position where it balances with the spring force of the compression spring 37. The control hydraulic pressure that is controlled to a predetermined pressure flowing in from the input port 33 is reduced, and a control hydraulic pressure that decreases as the control current from the control device increases is generated in the output port 38. The output port 38 of the linear solenoid pressure regulating valve 32 is connected to the control port 39 of the amplification valve 27 and to the switching port 40 of the switching valve 29. The amplifying valve 27 includes a compression spring 41 in which the axial force generated by the control hydraulic pressure of the linear solenoid pressure regulating valve 32 that is supplied from the control port 39 and acts on the large-diameter end surface of the valve body 49 acts on the small-diameter end surface of the valve body 49. The output hydraulic pressure Pc corresponding to the control hydraulic pressure of the linear solenoid pressure regulating valve 32 is output, which is moved to a position where the spring force of the motor and the axial force due to the feedback hydraulic pressure are balanced, and the line pressure supplied to the input port 28 decreases as the control current increases. Then, it is supplied from the output port 42 to the input port 43 of the switching valve 29.
[0034]
When the valve body 45 is shifted to the right half position in the figure, the switching valve 29 communicates the input port 43 with the output port 44, and outputs the output hydraulic pressure Pc from the amplification valve 27 to the hydraulic drive unit of the first clutch C-1. When the valve body 45 is shifted to the left half position in the figure, the line pressure port 30 communicates with the output port 44, and the line pressure from the port D of the manual valve 25 is supplied to the hydraulic drive unit of the first clutch C-1. And the first clutch C-1 is maintained in the engaged state by the line pressure. Hydraulic servo apparatus that supplies and discharges hydraulic pressure to each hydraulic drive section of the first to third clutches C-1 to C-3 and the first brake B-1 that are engaged and disengaged to establish each of the six forward speeds. Since all the components 26 have the same configuration, in the hydraulic circuit for engaging and disengaging these engaging elements shown in FIG. 5, a hydraulic servo device 26 for supplying and discharging hydraulic pressure to each engaging element, an amplifying valve 27, a switching valve 29, and The linear solenoid pressure adjusting valve 32 is displayed with the same reference number. For the linear solenoid, a reference sign indicating each engagement element is added after SL, and SLC-1, SLC-2, SLC-3, SLB- 1 is displayed.
[0035]
As shown in FIG. 5, the port D of the manual valve 25 to which the line pressure in which the hydraulic pressure from the oil pump P is controlled to a predetermined pressure by the pressure control valve 47 is supplied to the first clutch C via the cut-off valve 50. -1 is connected in parallel to the input port 28 of the amplification valve 27 of the hydraulic servo device 26 and the line pressure port 30 of the switching valve 29. The control port of the cut-off valve 50 is supplied with the hydraulic pressure supplied to the second clutch C-2 and the hydraulic pressure supplied to the third clutch C-3 or the first brake B-1, and the second clutch C-2. When the hydraulic pressure supplied to the third clutch C-3 or the first brake B-1 is high, the hydraulic pressure from the port D to the hydraulic servo device 26 for the first clutch C-1 is high. Shut off the supply. The port D of the manual valve 25 is connected in parallel to the input port 28 of the amplification valve 27 of the hydraulic servo device 26 for the second clutch C-2 and the line pressure port 30 of the switching valve 29 via the cutoff valve 51. Yes. The control port of the cut-off valve 51 is supplied with the hydraulic pressure supplied to the second clutch C-2 and the output hydraulic pressure of the solenoid valve 48, and when the hydraulic pressure supplied to the second clutch C-2 is high, or the solenoid valve When the output hydraulic pressure of 48 is high, supply of hydraulic pressure from the port D to the hydraulic servo device 26 for the second clutch C-2 is permitted.
[0036]
The switching valve 29 of the servo device 26 for the third clutch C-3 and the first brake B-1 is energized by the solenoid SLC-3 or SLB-1 of the linear solenoid pressure regulating valve 32 to switch from the output port 38 to the switching port 40. Valves 52 and 53 that are communicated with each other when the pressure supplied to the valve is lowered are added. The cut-off valve 54 to which the line pressure is supplied from the pressure control valve 47 is connected in parallel to the valves 52 and 53, and the valve 53 is connected to the input port 28 of the amplification valve 27 of the hydraulic servo device 26 for the third clutch C-3 and The valve 52 is connected in parallel to the line pressure port 30 of the switching valve 29, and the valve 52 is connected in parallel to the input port 28 of the amplification valve 27 of the hydraulic servo device 26 for the first brake B-1 and the line pressure port 30 of the switching valve 29. Has been. 55 is a cutoff connected between the output port 38 of the linear solenoid pressure regulating valve 32 having the solenoid SLC-3, the control port 39 of the amplification valve 27 for the third clutch C-3, and the switching port 40 of the switching valve 29. The hydraulic pressure supplied to the second clutch C-2 and the hydraulic pressure supplied to the first brake B-1 are supplied to the control port of the cutoff valve 55, and the second clutch C-2 and the first brake When the hydraulic pressure supplied to B-1 is high, the supply of hydraulic pressure from the output port 38 to the control port 39 and the switching port 40 is cut off. A cutoff 56 is connected between the output port 38 of the linear solenoid pressure regulating valve 32 having the solenoid SLB-1 and the control port 39 of the amplification valve 27 for the first brake B-1 and the switching port 40 of the switching valve 29. When the output side of the cut-off valve 55 is connected to the control port of the cut-off valve 56 and the output side hydraulic pressure of the cut-off valve 55 is high, the hydraulic pressure from the output port 38 to the control port 39 and the switching port 40 Shut off the supply. Reference numeral 57 denotes a valve for transmitting the hydraulic pressures of the third clutch C-3 and the first brake B-1 to the control port of the cut-off valve 50 without communicating with each other.
[0037]
The control device for the automatic transmission will be described based on the block diagram shown in FIG. The control device 60 with a built-in CPU includes an engine-side rotation speed sensor 61 that detects the engine-side rotation speed Ne of the torque converter 11 to which the engine rotation is transmitted, and an input shaft rotation speed sensor that detects the rotation speed Ni of the input shaft 20. 62, an output shaft rotational speed sensor 63 for detecting the rotational speed Nv of the output shaft 21, and a range position sensor 64 for transmitting detection signals Dr and Nr when the manual valve 25 is shifted to the forward travel range D and the neutral range N. Each detection signal is input from a throttle opening sensor 65 or the like that detects the accelerator depression amount Ss, and a control current that is a control signal is output to the linear solenoid pressure regulating valve 32 of each hydraulic servo device 26 based on these detection signals. In addition, when the first clutch C-1 that is always engaged when the low speed stage is established is abnormally engaged, the vehicle travels forward from the neutral range N. When a command to switch to any one of the high speed stages in the row range D is sent, the control program 70 shown in FIG. 7 is stored that prevents the engine from being driven by the drive wheels at an excessive speed. ing.
[0038]
Next, the operation of the embodiment of the control device for the automatic transmission according to the present invention will be described. When the manual valve 25 is shifted to the neutral range N during traveling at the fifth speed, the solenoids SLC-1 to SLC-3 and SLB-1 are turned on, and the first and second clutches C- The supply of hydraulic pressure to the hydraulic servo device 26 for 1 and C-2 is cut off, and the first and second clutches C-1 and C-2 are released. When the solenoids SLC-3 and SLB-1 are turned on, the output hydraulic pressure Pc from the hydraulic servo device 26 for the third clutch C-3 and the first brake B-1 becomes zero, and the third clutch C-1 and the first brake B -1 is released and becomes neutral.
[0039]
Thereafter, the control device 60 is returned to the forward travel range D, and when a command to switch from the neutral range N to any one of the high speed stages in the forward travel range D, for example, the fifth speed stage, is sent, the control device 60 is in a normal state. ND control which is a gradual increase control for gradually increasing the hydraulic pressure supplied from the hydraulic servo device 26 to the second clutch C-2 is performed, the solenoid SLC-3 is turned off, the third clutch C-3 is engaged, The fifth gear is established by engaging the two clutches C-2 so that the engagement shock does not occur. When the solenoid valve 48 is turned on for a predetermined time in the ND control, the hydraulic pressure is supplied to the control port and the second cut-off valve 51 is shifted, and the hydraulic pressure is transferred from the port D to the hydraulic servo device for the second clutch C-2. 26. Then, as shown in FIG. 8, the control device 60 controls the solenoid SLC-2 of the hydraulic servo device 26 for the second clutch C-2 in order to quickly close the gap between the clutch plates of the second clutch C-2. After the supplied control current I is reduced in pulses and the initial hydraulic pressure is supplied to the hydraulic drive of the second clutch C-2, the control current is slightly increased so that the second clutch C-2 The stroke pressure necessary to move the piston of the hydraulic drive unit to the position where slippage is started by contact is supplied, then the control current I is gradually decreased to gradually increase the output hydraulic pressure Pc, and the second clutch C-2 is shocked. Engage so as not to occur. Thereafter, the control current is reduced to 0, and the second clutch C-2 is maintained in the engaged state by the line pressure.
[0040]
At this time, in the case where the hydraulic servo device 26 for the first clutch C-1 has a failure in which the output hydraulic pressure is sent and the clutch C-1 is abnormally engaged even when the solenoid SLC-1 is turned on, the manual valve 25 Is shifted to the forward travel range D, the hydraulic pressure is supplied from the port D to the hydraulic servo device 26 for the first clutch C-1, the first clutch C-1 is abnormally engaged, and the second clutch C-2. The third speed is established until the is engaged. When the output hydraulic pressure Pc from the hydraulic servo device 26 for the second clutch C-2 rises and the second clutch C-2 is engaged, the fifth speed is established. The hydraulic pressure supplied to the second and third clutches C-2 and C-3 is supplied to the control port, the first cutoff valve 50 is shifted, and the hydraulic servo device 26 for the first clutch C-1 is connected to the port D. And the first clutch C-1 is released. As described above, the cutoff valve 50 is configured such that when the hydraulic pressure is applied to the hydraulic servo device 26 for the second clutch C-2 and the second clutch C-2 is engaged, the hydraulic servo for the first clutch C-1 is engaged. The hydraulic pressure applied to the device 26 is shut off.
[0041]
In ND control, when the solenoid SLC-3 is turned off, the increase in the output hydraulic pressure of the hydraulic servo device 26 for the third clutch C-3 is the hydraulic pressure for the first and second clutches C-1 and C-2. When the output hydraulic pressure of the servo device 26 is delayed for some reason, the hydraulic pressure supplied to the first and second clutches C-1 and C-2 is supplied to the control port, and the cutoff valve 54 is shifted. Since the supply of the line pressure to the hydraulic servo device 26 for the third clutch C-3 and the first brake B-1 is cut off, the third clutch C-3 is not engaged. Since the hydraulic pressure is not supplied from the third clutch C-3 to the control port, the first cutoff valve 50 does not disconnect the hydraulic servo device 26 for the first clutch C-1 from the port D, and the first clutch C- 1 may be engaged to establish the fourth speed.
[0042]
When the control device 60 recognizes from the detection signals Nr and Dr from the range position sensor 64 that the manual valve 25 has been shifted to the neutral range N and then returned to the forward travel range D while traveling at the fifth speed stage. In order to prevent the transmission mechanism 12 from shifting down to the low speed and causing the engine to rotate at the overspeed, a command for switching the transmission mechanism 12 from the neutral N to the fifth speed is sent. Execute. In step 71, it is determined whether or not the rotational speed Nv of the output shaft 21 is greater than a predetermined high speed rotational speed RH. The predetermined high-speed rotational speed RH is an input corresponding to the engine rotational speed limit in the Nv-Ni plane in which the horizontal axis shown in FIG. 9 is the rotational speed Nv of the output shaft 21 and the vertical axis is the rotational speed Ni of the input shaft 20. This is the rotational speed of the output shaft 21 at the intersection of the straight line indicating the rotational speed Rev limit of the shaft 20 and the straight line 4th indicating the fourth speed stage. The fourth gear is a gear that may be established when ND control is performed to switch to the fifth gear commanded from neutral N. If the rotational speed Nv is greater than the predetermined high speed rotational speed RH and the rotational speeds Nv and Ni are in the region (4) of the Nv-Ni plane, wait until the rotational speed Nv becomes smaller than the predetermined high speed rotational speed RH in step 72 To do. As a result, even if the fourth speed is established by ND control, the engine is not rotated more than the limit rotational speed by the drive wheels.
[0043]
If the rotational speed Nv is smaller than the predetermined high speed rotational speed RH, it is determined in step 73 whether the rotational speed Nv of the output shaft 21 is smaller than the predetermined low speed rotational speed RL. If the rotational speed Nv is smaller than the predetermined low speed rotational speed RL and is in the region (3) on the Nv-Ni plane, the engine will be driven by the drive wheels beyond the limit rotational speed even if the third and fourth speed stages are established. Since it is not rotated, ND control is performed in step 74 as usual. The predetermined low-speed rotational speed RL is the value of the output shaft 21 at the intersection of the straight line indicating the rotational speed Rev limit of the input shaft 20 corresponding to the engine rotational speed limit and the straight line 3rd indicating the third speed stage in the Nv-Ni plane. The number of revolutions.
[0044]
When the rotational speed Nv is larger than the predetermined low speed rotational speed RL, the rotational speed Ni of the input shaft 20 detected by the input shaft rotational speed sensor 62 in step 75 is the rotation of the output shaft 21 detected by the output shaft rotational speed sensor 63. It is determined whether or not it is smaller than the ideal rotational speed (Nv × Gr5) multiplied by the gear ratio Gr5 of the fifth speed, which is the gear stage commanded to the number Nv. If it is small and falls within the area (2) of the Nv-Ni plane, as shown in FIG. 8, the fifth speed is commanded in step 76 to perform ND control, and the hydraulic pressure for the second clutch C-2 The output hydraulic pressure Pc of the servo device 26 gradually increases and the second clutch C-2 is gradually engaged. In step 77, the rotational speed Ni of the input shaft 20 is increased until it is substantially equal to the ideal rotational speed (Nv × Gr5) of the input shaft 20 obtained by multiplying the rotational speed Nv of the output shaft 21 by the fifth gear ratio Gr5. Is determined, the ND control is interrupted at step 78, the control current applied to the solenoid SLC-2 is reduced to 0, the second clutch C-2 is engaged by the line pressure, and the fifth speed stage is established. Is established.
[0045]
When the rotational speed Ni of the input shaft 20 is larger than the ideal rotational speed (Nv × Gr5) and is in the area (1) of the Nv-Ni plane, the third speed stage is being executed during the ND control commanding the fifth speed stage. Is established, the estimated rotational speed (Nv × Gr3) obtained by multiplying the rotational speed Nv of the output shaft 21 by the gear ratio Gr3 of the third speed exceeds the rotational speed Rev limit of the input shaft 20 corresponding to the engine rotational speed limit. Since there is a possibility that the engine is rotated at an excessive speed by the drive wheel, the engine torque limit control is performed in step 79 to limit the engine output torque, and in step 80 the rotational speed Ni of the input shaft 20 is If it is determined that the rotational speed Nv of the output shaft 21 has decreased to approximately equal to the ideal rotational speed (Nv × Gr5) of the input shaft 20 obtained by multiplying the gear ratio Gr5 of the fifth speed stage, in step 81, ND Without control, the hydraulic pressure is supplied to and discharged from the second clutch C-2 as shown in FIG. The output hydraulic pressure of the hydraulic servo device 26 is steeply increased, and the hydraulic steep control for quickly engaging the second clutch C-2 is performed.
[0046]
When the manual valve 25 is shifted to the neutral range after traveling at the sixth speed and then returned to the forward travel range D, the solenoid SLB-1 is turned off by the ND control commanding the sixth speed. The second speed is established between the time when the first brake B-1 is engaged and the output hydraulic pressure Pc of the hydraulic servo device 26 for the solenoid SLC-2 is gradually increased until the second clutch C-2 is engaged. Since other than that is the same as the case of the fifth gear described above, detailed description thereof will be omitted.
[0047]
In the above-described embodiment, there is a possibility that it is established when a straight line indicating the rotational speed Rev limit of the input shaft 20 corresponding to the engine rotational speed limit and ND control for switching to the fifth speed commanded from the neutral N are performed. The rotation speed of the output shaft 21 at the intersection point with the straight line 4th indicating the fourth speed stage is set to the predetermined high speed rotation speed RH. If the gear ratio of the possible gear is smaller, the rotation speed of the output shaft 21 at the intersection of the straight line indicating the rotation speed Rev limit and the straight line indicating the commanded gear shift is set to the predetermined high speed rotation speed RH. .
[0048]
Further, in the above embodiment, it is determined whether or not the rotational speed Ni of the input shaft 20 is larger than the ideal rotational speed obtained by multiplying the rotational speed Nv of the output shaft 21 by the commanded fifth gear ratio Gr5. However, the ratio between the rotational speed Ni of the input shaft 20 detected by the input shaft rotational speed sensor 62 and the output shaft rotational speed sensor 63 and the rotational speed Nv of the output shaft 21 is calculated as an actual gear ratio. It may be determined whether or not the gear ratio is larger than the commanded fifth gear stage gear ratio Gr5.
[Brief description of the drawings]
FIG. 1 is a skeleton diagram of an embodiment of an automatic transmission controlled by a control device for an automatic transmission according to the present invention.
FIG. 2 is an engagement table of clutches and brakes at each gear stage of the automatic transmission.
FIG. 3 is a velocity diagram showing a rotation speed ratio of each element of the planetary gear at each shift stage of the automatic transmission.
FIG. 4 is a view showing a hydraulic servo device that engages and disengages a first clutch C-1.
FIG. 5 is a hydraulic circuit diagram for supplying and discharging hydraulic pressure to a hydraulic servo device.
FIG. 6 is a block diagram showing a control device for an automatic transmission.
FIG. 7 is a flowchart of a control program.
FIG. 8 is a diagram showing an engaged state of each clutch when ND control is performed when shifting from a neutral range to a forward travel range.
FIG. 9 is a diagram in which regions of both rotational speeds are divided in order to perform control suitable for the rotational speeds of the output shaft and the input shaft.
FIG. 10 is a diagram showing an engaged state of each clutch when ND control is not performed when shifting from the neutral range to the forward travel range.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 10 ... Automatic transmission, 11 ... Torque converter (fluid transmission device), 12 ... Transmission mechanism, 19 ... Lock-up clutch, 20 ... Input shaft, 21 ... Output shaft, 25 ... Manual valve, 26 ... Hydraulic servo device, 27 ... Amplification valve, 29 ... Switching valve, 32 ... Linear solenoid pressure regulator, SLC-1 to SLC-3, SLB-1, SLB -2 ... Linear solenoid, 36, 49 ... Valve body, 47 ... Pressure control valve, 48 ... Solenoid valve, 51, 54-56 ... Cut-off valve, 52, 53 ... Valve, 60 ... Control device, 61 ... Engine side rotational speed sensor, 62 ... Input shaft rotational speed sensor, 63 ... Output shaft rotational speed sensor, 64 ... Range position sensor, 65 ...・ Throttle opening sensor, C-1 ... 1st clutch, C-2 ... 2nd clutch, C-3 · Third clutch, B-1, ... first brake, B-2 ... second brake, 70 ... control program.

Claims

A speed change mechanism that shifts the rotation of the input shaft and outputs it to the output shaft, a plurality of engagement elements that are engaged and disengaged to establish each gear position of the speed change mechanism, and an electric signal supplied to the solenoid A plurality of hydraulic servo devices that supply and discharge output hydraulic pressure to engage and disengage the engaging elements respectively, an input shaft rotational speed detection device that detects the rotational speed of the input shaft and the rotational speed of the output shaft of the transmission mechanism, and an output In the automatic transmission control device including the shaft rotation speed detection device, the plurality of engagement elements are a high-speed stage engagement element that is always engaged when the high-speed stage of the transmission mechanism is established, and a constant engagement when the low-speed stage is established. A low-speed stage engagement element to be combined, a third engagement element that is engaged together with the high-speed stage engagement element or the low-speed stage engagement element, and establishes the high-speed stage or the low-speed stage, and Including four engaging elements, said high to establish said high speed stage By Dangakarigo element and the third engagement element and the fourth engagement element is engaged, it has a blocking means for blocking the supply of hydraulic pressure to the low speed stage engaging elements, wherein the control device, running Means for determining whether the rotational speed of the output shaft is smaller than a predetermined low-speed rotational speed when a command to switch the speed change mechanism from neutral to any one of the high-speed speeds is sent, Means for gradually increasing the hydraulic pressure supplied from the hydraulic servo device to the high-speed stage engaging element to switch the speed change mechanism from neutral to the commanded gear stage; and, if large, the hydraulic pressure gradually increasing control. Means for steeply increasing the output hydraulic pressure of the hydraulic servo device that supplies or discharges the hydraulic pressure to or from the high-speed stage engaging element without interrupting or performing the hydraulic steep control for quickly engaging the high-speed stage engaging element. Specially provided Control apparatus for an automatic transmission according to.

The predetermined low-speed rotation speed is equal to the engine speed limit when multiplied by the low-speed gear ratio that is established in the middle of the hydraulic gradual increase control when the low-speed engagement element is abnormally engaged. 2. The control device for an automatic transmission according to claim 1, wherein the number of rotations is.

When the rotational speed of the output shaft is greater than the predetermined low-speed rotational speed, the control device is configured such that the rotational speed of the input shaft is obtained by multiplying the rotational speed of the output shaft by the gear ratio of the commanded gear. Means for determining whether or not the actual gear ratio between the rotational speed of the input shaft and the rotational speed of the output shaft is greater than the gear ratio of the commanded gear, A hydraulic servo apparatus that supplies and discharges hydraulic pressure to and from the high-speed engagement element without performing the gradual increase control when the rotational speed of the input shaft becomes substantially equal to the ideal rotational speed by performing torque limit control of the engine 3. The control apparatus for an automatic transmission according to claim 1 or 2, further comprising means for steeply increasing the output hydraulic pressure so as to perform hydraulic steep control for quickly engaging the high-speed stage engaging element.

When the rotational speed of the output shaft is larger than the predetermined low-speed rotational speed, the control device has a rotational speed or actual gear ratio of the input shaft smaller than the ideal rotational speed or the gear ratio of the commanded gear. In this case, when the rotational speed of the input shaft becomes substantially equal to the ideal rotational speed, the output hydraulic pressure of the hydraulic servo device that interrupts the gradual increase control and supplies / discharges hydraulic pressure to / from the high-speed engaging element is increased steeply. 5. The control apparatus for an automatic transmission according to claim 3 or 4, further comprising means for performing hydraulic steep control for quickly engaging the high speed engagement element.

The control device determines whether or not the rotational speed of the output shaft is greater than a predetermined high-speed rotational speed, and if so, waits until the rotational speed of the output shaft decreases below the predetermined high-speed rotational speed. The control device for an automatic transmission according to any one of claims 1 to 4, further comprising:

The predetermined high-speed rotational speed is multiplied by a gear ratio of a gear stage that may be established or a gear ratio of a commanded gear stage that may be established when hydraulic gradual increase control for switching to a high-speed gear stage commanded from neutral is performed. 6. The control device for an automatic transmission according to claim 5, wherein the rotational speed of the output shaft is equal to the limit rotational speed of the engine.

Cut-off that cuts off the hydraulic pressure applied to the hydraulic servo device for the low speed stage engaging element when the hydraulic pressure is applied to the hydraulic servo apparatus for the high speed stage engaging element and the high speed stage engaging element is engaged. The control device for an automatic transmission according to any one of claims 1 to 6, further comprising a valve.