JP4253891B2 - Powertrain control device - Google Patents

Description

【０１１９】
【数２】

【０１２０】
【数３】

【０１２１】
【数４】

次に、図１２に示すように、上記モード切換ポイント（ウ）は、ローモードＬＦ及びハイモードＨＦのいずれにおいても、同じトロイダルレシオ（Ｒｔｍ）で同じユニットレシオ（Ｒｕｍ）が得られる唯一のポイントであり、したがってこのポイント（ウ）でモードの切換えを行なうことにより、切換え前後でユニットレシオの著しい変動のない、したがって切換えショックの発生しない円滑なモードの切換えが実現する。そして、この切換ポイントトロイダルレシオ（Ｒｔｍ）を実現させる切換ポイントパルス数（Ｎｍ：トルクゼロの場合）ないし切換ポイントスリーブ位置（Ｓｍ：トルクゼロの場合）といった各種の物理量が、前述の図８、図９及び図１１に示すように、理論的にただ一点対応して定まっている。なお、上記の切換ポイントトロイダルレシオ（Ｒｔｍ）を実現させる切換ポイントパルス数ないし切換ポイントスリーブ位置がトルクに応じて変化することは前述の通りである。
【０１２２】
それゆえ、理論上は、実トロイダルレシオが切換ポイントトロイダルレシオ（Ｒｔｍ）に到達し、したがって、パルス数ないしスリーブ位置が切換ポイントパルス数（Ｎｍ）ないし切換ポイントスリーブ位置（Ｓｍ）に到達したときに、そのパルス数ないしスリーブ位置の制御を停止すれば、実トロイダルレシオが上記切換ポイントトロイダルレシオ（Ｒｔｍ）において安定することになるのであるが、このモード切換ポイント（ウ）の近傍においては、該切換ポイントに接近するようにパワーローラ２３，３３が連続的に傾転しており、したがって、現実には、その連続的な傾転運動による慣性等が働く結果、実トロイダルレシオが切換ポイントトロイダルレシオ（Ｒｔｍ）に到達してから、パルス数ないしスリーブ位置の制御を停止したのでは、パワーローラ２３，３３が傾転し過ぎ、その結果、実トロイダルレシオが切換ポイントトロイダルレシオ（Ｒｔｍ）を越えてより減速側（ローモードＬＦからハイモードＨＦ、及びハイモードＨＦからローモードＬＦのいずれのモード切換えにおいても、トロイダルレシオは、その切換え前は、増速側から減速側に変化している）に行き過ぎることになるのである（図１２に鎖線で示す領域に進入する）。
【０１２３】
そして、ローラ２３，３３の傾転運動の慣性は伝達トルクが大きくなるに従ってより大きくなるから、上記のローラ２３，３３の行き過ぎ量も伝達トルクに応じて大きくなる。したがって、モード切換時、特にその開始時に、そのようなパワーローラ２３，３３の行き過ぎに起因する実レシオの切換ポイントから減速側へのずれを考慮しないと、モード切換動作の進行に伴って著しいトロイダルレシオないしユニットレシオの変動が起こり、不快なショックが発生することになる。
【０１２４】
この実施の形態に係る無段変速機構２０，３０を搭載したパワートレイン１０においては、このような不具合を解消するための対策もまた講じられている。
【０１２５】
次に、上記の前進ローモードＬＦと前進ハイモードＨＦとの切換時の具体的制御動作を詳しく説明する。この制御は、基本的に、モード切換ポイント（ウ）近傍における制御であり、各モード達成用の摩擦要素としてのロークラッチ６０とハイクラッチ７０との掛け替えが行なわれる。すなわち、ローモードＬＦからハイモードＨＦへの切換えであれば、ロークラッチ６０を切断し、ハイクラッチ７０を締結する。逆に、ハイモードＨＦからローモードＬＦへの切換えであれば、ハイクラッチ７０を切断し、ロークラッチ６０を締結する。そして、このクラッチ６０，７０の掛け替え動作は、実トロイダルレシオが切換ポイントトロイダルレシオ（Ｒｔｍ）に到達することをもって開始され、その掛け替え動作中は、実トロイダルレシオが切換ポイントトロイダルレシオ（Ｒｔｍ）に一定に保持されるように、前進用ステップモータ２５１に対するパルス数の制御が行われる。
【０１２６】
一方、このようなモードの切換時以外の通常時は、基本的に、図１３〜図１５に示すように、前進ローモードＬＦ、前進ハイモードＨＦ、及び後退ローモードＬＲのそれぞれにおいて、車速（Ｖ）やスロットル開度（ＴＶＯ）等の車両の走行状態をパラメータとして予め設定された変速線図に基づく変速比（トロイダルレシオ及びユニットレシオ）のフィードバック制御が行なわれている。
【０１２７】
この変速比制御は、特に図１４に示すように、まず実車速（Ｖ）と実スロットル開度（ＴＶＯ）とを上記変速線図にあてはめて目標エンジン回転数（Ｎｅｏ）を求め、次にこの目標エンジン回転数（Ｎｅｏ）と実車速（Ｖ）とから目標ユニットレシオ（Ｒｕｏ）を算出して、さらにこの目標ユニットレシオ（Ｒｕｏ）が得られる目標トロイダルレシオ（Ｒｔｏ）を設定したのち、この目標トロイダルレシオ（Ｒｔｏ）が実現するように、ステップモータ２５１，２５２に対するパルス制御（スリーブ位置制御）を介してトロイダルレシオをフィードバック制御するものである。
【０１２８】
なお、各変速線図には、モード切換ポイントユニットレシオ（Ｒｕｍ）の傾きを有するモード切換ライン（Ｍ）が表されている。
【０１２９】
図１６のタイムチャートを参照してさらに説明する。なお、このタイムチャートは、一例として、例えばアクセルペダルが踏み込まれ続けて、車速が増加していき、その結果、正駆動状態でローモードＬＦからハイモードＨＦへの切換えが起こる場合を示している。
【０１３０】
まず時点ｔ１までの期間中、及び時点ｔ４以降の期間中は、上記の走行状態及び変速線図に基づく通常の変速比のフィードバック（Ｆ／Ｂ）制御が行なわれ、そして、これらの間の時点ｔ１から時点ｔ４までの期間中にモードの切換制御が実行される。
【０１３１】
まず時点ｔ１までの期間中はローモードＬＦが達成されており、したがって、オンオフソレノイドバルブ２１３がＯＮとされた状態で、ロークラッチ用デューティソレノイドバルブ２７１に対するデューティ率ＤＬが１００％（ＤＬ１）とされ、ハイクラッチ用デューティソレノイドバルブ２７２に対するデューティ率ＤＨが０％（ＤＨ１）とされている。これにより、ロークラッチ６０の油圧室にはロークラッチ圧（ＥＬ）としてライン圧がそのまま供給され、該ロークラッチ６０が完全締結状態とされている一方で、ハイクラッチ７０の油圧室にはハイクラッチ圧（ＥＨ）が供給されておらず、該ハイクラッチ７０が完全解放状態とされている。
【０１３２】
そして、この時点ｔ１までの期間中は、変速比のフィードバック制御により、目標ユニットレシオ（Ｒｕｏ）が増速側に変化するように設定されていき、したがって図中鎖線で示すように、目標トロイダルレシオ（Ｒｔｏ）が減速側に変化するように設定されていって、この目標トロイダルレシオ（Ｒｔｏ）が実現するようにトロイダルレシオ（Ｒｔ）がフィードバック制御される結果、図中実線で示すように、実トロイダルレシオ（Ｒｔｒ）が目標トロイダルレシオ（Ｒｔｏ）に追随するように減速側に変化していく。このとき、三層弁２２０についてのパルス数（Ｎ）ないしスリーブ位置（Ｓ）は共にマイナス側に変化していく。
【０１３３】
そして、符号カで示すように、実トロイダルレシオ（Ｒｔｒ）が、切換ポイントトロイダルレシオ（Ｒｔｍ）より手前の増速側の所定トロイダルレシオ（Ｒｔａ）に到達した時点ｔ１において、変速比のフィードバック制御が停止されると共に、パルス数（Ｎ）ないしスリーブ位置（Ｓ）が、符号キで示すように、その時点ｔ１における所定パルス数（Ｎａ）及び所定スリーブ位置（Ｓａ）に固定される。このパルス数（Ｎ）等の固定は、時点ｔ１から所定時間Ｔａが経過するまで続けられる。
【０１３４】
ここで、上記所定トロイダルレシオ（Ｒｔａ）は、図１７に示すように、ｔ１の時点で変速機構２０，３０を通過するトルク（Ｔｒ１）に応じて補正され、該トルク（Ｔｒ１）がゼロでパワーローラ２３，３３の傾転慣性が生じないときは、理論値としての切換ポイントトロイダルレシオ（Ｒｔｍ）とされるが、上記トルク（Ｔｒ１）が大きくなり、したがってパワーローラ２３，３３の傾転慣性が大きくなるに伴って、より増速側の値に設定される。これにより、パワーローラ２３，３３は慣性により減速側への傾転を続けたのち停止し、その結果、符号クで示すように、時点ｔ１よりのちの時点ｔｘにおいて、実トロイダルレシオ（Ｒｔｒ）が切換ポイントトロイダルレシオ（Ｒｔｍ）に到達して、該レシオ（Ｒｔｍ）で安定する。そして、少なくとも、モードの切換えが終了して再び通常の変速比のフィードバック制御が再開される時点ｔ４が過ぎるまでは、実トロイダルレシオ（Ｒｔｒ）が上記切換ポイントトロイダルレシオ（Ｒｔｍ）に安定に固定するように、前進用ステップモータ２５１へのパルス制御が行われる。これにより、著しい変速比変動及び切換えショックの発生しない円滑なクラッチ６０，７０の掛け替えが実現することになる。
【０１３５】
これに対し、実トロイダルレシオ（Ｒｔｒ）が切換ポイントトロイダルレシオ（Ｒｔｍ）に到達するまで、符号ケで示すように、パルス数（Ｎ）ないしスリーブ位置（Ｓ）をマイナス側にフィードバック制御し続け、そして、実トロイダルレシオ（Ｒｔｒ）が切換ポイントトロイダルレシオ（Ｒｔｍ）に到達してから、その制御を停止したのでは、符号コで示すように、実トロイダルレシオ（Ｒｔｒ）が切換ポイントトロイダルレシオ（Ｒｔｍ）を越えてさらに減速側に行き過ぎ、その結果、変速比が切換ポイント変速比からずれて、クラッチ６０，７０の掛け替え時に不快なショックが発生したり、又は、符号サで示すように、エンジン回転数（Ｎｅ）が不必要に上昇したりするのである。
【０１３６】
なお、実トロイダルレシオ（Ｒｔｒ）が所定トロイダルレシオ（Ｒｔａ）に到達した時点ｔ１における上記所定パルス数（Ｎａ）及び所定スリーブ位置（Ｓａ）は、すなわち、上記伝達トルク（Ｔｒ１）のもとで所定トロイダルレシオ（Ｒｔａ）を実現させることのできるパルス数（Ｎ）及びスリーブ位置（Ｓ）であるということができるが、さらに、前述の図１１で説明したトラクション力に起因して発生する目標値からのずれ現象も加味されている。
【０１３７】
さらに、このｔ１の時点で、ハイモードＨＦに切り換わった後に時点ｔ４から再開される変速比のフィードバック制御で用いられるその再開始時の目標トロイダルレシオ（Ｒｔｂ）が算出される。その場合にも、前述の図１１で説明したトラクション力に起因して発生する目標値からのパルス数ないしスリーブ位置のずれ現象が加味される。
【０１３８】
一方、実トロイダルレシオ（Ｒｔｒ）が所定トロイダルレシオ（Ｒｔａ）に到達した時点ｔ１で、ロークラッチ用デューティ率ＤＬが０％（ＤＬ２）とされ、ハイクラッチ用デューティ率ＤＨが所定デューティ率（ＤＨ２）に設定される。これにより、ロークラッチ６０が開放され始める一方、ハイクラッチ７０が締結され始める。すなわち、実トロイダルレシオ（Ｒｔｒ）が所定トロイダルレシオ（Ｒｔａ）に到達したことをもってクラッチ６０，７０の掛け替え動作が開始されるのである。
【０１３９】
ここで、パルス数（Ｎ）ないしスリーブ位置（Ｓ）が時点ｔ１における所定パルス数（Ｎａ）及び所定スリーブ位置（Ｓａ）に固定され続ける上記所定時間Ｔａは、実質的に、このクラッチ６０，７０の掛け替え動作の開始に伴い、ロー、ハイの両クラッチ６０，７０ともが、符号タで示すように、締結状態となるのに要する時間とされている。そして、ロー、ハイの両クラッチ６０，７０ともが締結状態となったときには、パルス数（Ｎ）ないしスリーブ位置（Ｓ）の如何に拘らず、トロイダルレシオは切換えポイントトロイダルレシオ（Ｒｔｍ）に堅持される。したがって、ｔ１の時点で実トロイダルレシオ（Ｒｔｒ）が所定トロイダルレシオ（Ｒｔａ）に到達さえすれば、あとはクラッチ６０，７０の掛け替え動作の開始によってトロイダルレシオが切換えポイントトロイダルレシオ（Ｒｔｍ）に安定的に保持されるから、もはやパルス数（Ｎ）等のフィードバック制御は行う意味がなく、それゆえｔ１の時点で該フィードバック制御が停止されて、パルス数（Ｎ）等が上記所定パルス数（Ｎａ）等に固定されるのである。
【０１４０】
なお、図１８に示すように、上記所定時間Ｔａは、油温が高いほど短くなるように設定される。これは、油温が高いときは作動油ないし潤滑油の粘性が低く、締結側摩擦要素（この場合はハイクラッチ７０）の締結動作が応答性よく進行すると共に、また実トロイダルレシオ（Ｒｔｒ）が切換えポイントトロイダルレシオ（Ｒｔｍ）に速やかに到達するからである。これにより、このモード切換えに要する制御時間の短縮化が図られる。
【０１４１】
そして、上記所定時間Ｔａが経過した時点ｔ２以降は、ハイクラッチ用デューティ率ＤＨが１００％のデューティ率（ＤＨ３）に向けてさらに高められる。これにより、ロー、ハイの両クラッチ６０，７０ともが締結状態となった状態から、ロークラッチ６０の開放動作及びハイクラッチ７０の締結動作がさらに進むことになる。すなわち、ローモードからハイモードへの移行が開始されるのである。
【０１４２】
このとき、時点ｔ２から時点ｔ３までの第２所定時間Ｔｂが経過する間に、パルス数（Ｎ）ないしスリーブ位置（Ｓ）のフィードフォワード（Ｆ／Ｆ）制御が実行される。すなわち、時点ｔ１におけるトルク（Ｔｒ１）をモード切換え前のローモードＬＦでのトルクとし、時点ｔ２におけるトルク（Ｔｒ２）をモード切換え後のハイモードＨＦでのトルクとして、前述の図１１を参照して説明したように、両モードＬＦ，ＨＦで切換ポイントトロイダルレシオ（Ｒｔｍ）が保持実現するように、そのＲｔｍ実現パルス数Ｎ１，Ｎ２の偏差Δｎだけステップモータ２５１に対するパルス数（Ｎ）を修正するのである。ここで、Ｒｔｍ実現パルス数の算出には、前述の特性近似式ｆ１〜ｆ４が用いられ、特に、この場合は、ローモードＬＦ、正駆動状態でのＲｔｍ実現パルス数Ｎ１の算出に特性近似式ｆ１が、またハイモードＨＦ、正駆動状態でのＲｔｍ実現パルス数Ｎ２の算出に特性近似式ｆ２が用いられる。
【０１４３】
これにより、符号チで示すように、パルス数（Ｎ）は第２所定時間Ｔｂの間にフィードフォワード制御によって応答性よく速やかに偏差Δｎだけ増速側にプラスされる。したがって、ロー、ハイの両クラッチ６０，７０ともが締結状態となった状態からハイモードへの移行が開始され、その結果、パワーローラ２３，３３が前述のトラクション力Ｔ２，Ｔ２によって増速方向にオフセットした位置から、反対方向に作用するトラクション力Ｔ１，Ｔ１によって減速方向にオフセットした位置へ移動することになっても、符号ツで示すように、著しいトロイダルレシオないしユニットレシオの変動が起こらず、不快なショックの発生が抑制されることになる。
【０１４４】
これに対し、符号テで示すように、パルス数（Ｎ）をハイモードへの移行の開始前と同じ値に固持したときには、パワーローラ２３，３３のオフセット方向の反転に起因する、符号トで示すようなトロイダルレシオの減速方向への変動を回避できず、その結果、変速比が切換ポイント変速比からずれて不快な切換ショックが発生したり、又は、符号ナで示すように、エンジン回転数（Ｎｅ）がここでもまた不必要に上昇したりするのである。
【０１４５】
なお、このパルス数の偏差Δｎは、換言すれば、現状の実トロイダルレシオである切換ポイントトロイダルレシオ（Ｒｔｍ）と、ハイモードＨＦに切り換わった後に時点ｔ４から再開される変速比のフィードバック制御で用いられる目標トロイダルレシオ（Ｒｔｂ）との間の差分（Δｒ）を修正するものである。
【０１４６】
一方、時点ｔ２から時点ｔ３までの第２所定時間Ｔｂが経過する間、エンジン１の点火時期（Ｉｇ）が所定量ΔＩｇだけリタードされて、入力トルクの低減が図られる。その場合に、上記リタード量ΔＩｇは、図１９に示すように、時点ｔ２における入力トルク（Ｔｒ２）に応じて補正され、該トルク（Ｔｒ２）が大きくなるほど、リタード量ΔＩｇが大きく、つまりトルクダウン量が大きくなるように設定される。これにより、ハイクラッチ７０の締結ショックが抑制されるばかりでなく、その締結動作の進行に伴うトルク変動が抑制され、したがってパワーローラ２３，３３のオフセット反転量も低減されて、モード切換えショックがより一層確実に低減されることになる。
【０１４７】
そして、第２所定時間Ｔｂが経過し、時点ｔ３に至った以降は、パルス数（Ｎ）のフィードバック制御は依然として停止される一方、オンオフソレノイドバルブ２１３がオフとされる。これにより、フェールセーフバルブ２１２のスプールが左側に移動して、ロークラッチライン２７４が該バルブ２１２のドレインポート２１２ｂに連通する。それゆえ、符号ハで示すように、ロークラッチ圧がこのドレインポート２１２ｂから速やかに排出され、ロークラッチ６０が早期に完全開放状態となって、切り換わり後のハイモードＨＦにおける変速比のフィードバック制御を早期に実行することが可能となる。
【０１４８】
これに対し、オンオフソレノイドバルブ２１３がオンのままであると、ロークラッチ圧がデューティソレノイドバルブ２７１のデューティ率により排出されることになり、符号ヒで示すように、遠心残圧等が残って、ハイモードＨＦが完全に実現するのに時間がかかることになる。
【０１４９】
そして、時点ｔ２からの所定値Ｔｃが経過した段階で、ハイクラッチ７０が完全締結され、ロークラッチ６０が完全解放されたハイモードＨＦに完全に切り換わり、上記トロイダルレシオ（Ｒｔｂ）を目標とする変速比のフィードバック制御が再開されることになる。
【０１５０】
次に、以上の制御動作を図２０以下のフローチャートに沿って説明する。まず、ステップＳ１からＳ１７までは時点ｔ１に至るまでの制御動作である。
【０１５１】
まずステップＳ１で、図７に示す各センサやスイッチ３０１〜３０８からの信号に基づき、現時点における車速、エンジン１のスロットル開度、選択されているレンジ、実トロイダルレシオ、油温等の各種の状態量を検出したうえで、ステップＳ２において入力トルクＴｒを推定する。この入力トルクＴｒの推定はエンジン回転数や吸入空気量等から周知の方法により求められる。次いでステップＳ３でエンジン点火時期タイミングＩｇを設定する。この点火時期タイミングＩｇもまたエンジン回転数や吸入空気量等から周知の方法により求められる。
【０１５２】
次いでステップＳ４で上記入力トルクＴｒに応じてモード切換判定レシオ（所定トロイダルレシオ）Ｒｔａを設定する。その場合に、モード切換判定レシオＲｔａは、前述の図１７に示すように、入力トルクＴｒが大きいほど小さい値（増速側の値）に求められる。
【０１５３】
次いでステップＳ５で実トロイダルレシオＲｔｒが上記モード切換判定レシオＲｔａよりも大きいか否かを判定する。その結果、大きくないとき、つまり実トロイダルレシオＲｔｒがモード切換判定レシオＲｔａより増速側であるときは、ステップＳ６に進んで、上記入力トルクＴｒの値を第１トルク値Ｔｒ１とする。
【０１５４】
次いでステップＳ７でトロイダルレシオの目標値Ｒｔｏをモード切換理論値（切換ポイントトロイダルレシオ）Ｒｔｍとし、ステップＳ８、Ｓ９でハイクラッチデューティ率ＤＨを第１デューティ率ＤＨ１（０％）とし、ロークラッチデューティ値ＤＬを第１デューティ率ＤＬ１（１００％）とし、さらにステップＳ１０でオンオフソレノイドバルブ２１３に対する指示値をオンとする。
【０１５５】
次いでステップＳ１１では、まず変速線図に基づいて目標ユニットレシオＲｕｏないし目標トロイダルレシオＲｔｏを設定したのち、該目標トロイダルレシオＲｔｏに対する実トロイダルレシオＲｔｒの偏差ΔＲｔを算出する。
【０１５６】
ステップＳ１２では、入力トルク、変速比、走行モード、及びレンジに応じて、ＰＩＤ制御の指数Ｑにおける比例項ゲイン（変速比制御におけるフィードバックゲイン）Ｇを設定する。この比例項ゲインＧは、トルクが大きいほど、またトロイダルレシオがＧＮレシオ（Ｒｔｎ）に近いほど小さい値に設定される。さらに、走行モードがローモードＬＦ，ＬＲであるときはハイモードＨＦであるときに比べて小さい値に設定され、またレンジが後退レンジＬｒであるときは前進レンジであるときに比べて小さい値に設定される。
【０１５７】
ステップＳ１３では、まず次式に比例項ゲインＧを代入してＰＩＤ制御指数Ｑを算出する。
【０１５８】
【数５】

ここで、Ｈは積分項ゲイン、及びＩは微分項ゲインである。
【０１５９】
次いで、予め図２２に示すように設定されたマップに基づいて、算出した指数Ｑに応じたステップモータ２５１，２５２に出力する制御信号のパルス数偏差ΔＮを求める。
【０１６０】
ここで、上記マップでは、指数Ｑの絶対値が大きくなるほどパルス数偏差ΔＮも大きくなるように設定されている。また、指数Ｑが正のとき（実トロイダルレシオが目標トロイダルレシオよりも大きいとき）には、パルス数偏差ΔＮを正として、パルス数Ｎを増加させ、これにより実トロイダルレシオを小さくし、逆に、指数Ｑが負のとき（実トロイダルレシオが目標トロイダルレシオよりも小さいとき）には、パルス数偏差ΔＮを負として、パルス数Ｎを減少させ、これにより実トロイダルレシオを大きくするように設定されている（図８参照）。
【０１６１】
そして、ステップＳ１４〜Ｓ１７において、ステップモータ２５１、デューティソレノイドバルブ２７１，２７２、オノフソレノイドバルブ２１３、及びエンジン１の点火時期をそれぞれ制御することにより、前述したような時点ｔ１までの制御動作が得られる。
【０１６２】
一方、ステップＳ５で実トロイダルレシオＲｔｒがモード切換判定レシオＲｔａに到達したときには、該ステップＳ５からステップＳ１８に進む。この場合のステップＳ１からＳ５及びＳ１８からＳ２４及びＳ１４からＳ１７までの動作は時点ｔ１から時点ｔ２に至るまでの制御動作である。
【０１６３】
ステップＳ１８で実トロイダルレシオＲｔｒがモード切換判定レシオＲｔａに到達した直後に限り、第１タイマーｔｉｍ１をセットする。そして、ステップＳ１９で該タイマーｔｉｍ１が第１所定期間Ｔａを過ぎたか否かが判定される。その結果、まだ第１所定期間Ｔａを過ぎていないとき、つまり時点ｔ２に至るまでは、ステップＳ２０で入力トルクＴｒの値を第２トルク値Ｔｒ２とする。
【０１６４】
また、ステップＳ２１でステップモータ２５１に対するパルス数偏差ΔＮとして０を代入し、次いでステップＳ２２、Ｓ２３でハイクラッチ用デューティ率ＤＨを所定の棚圧である第２デューティ率ＤＨ２とし、またロークラッチ用デューティ率ＤＬを０％である第２デューティ率ＤＬ２とする。さらにステップＳ２４でオンオフソレノイドバルブ２１３に対する指示値をオンとする。
【０１６５】
そして、ステップＳ１４〜Ｓ１７において、ステップモータ２５１、デューティソレノイドバルブ２７１，２７２、オノフソレノイドバルブ２１３、及びエンジン１の点火時期をそれぞれ制御することにより、前述したような時点ｔ１から所定時間Ｔａが経過する時点ｔ２までの制御動作が得られる。
【０１６６】
一方、ステップＳ１９で第１タイマーｔｉｍ１が第１所定期間Ｔａを過ぎたときは、該ステップＳ１９からＳ２５に進む。この場合のステップＳ１からＳ５及びＳ１８，Ｓ１９及びＳ２５からＳ３２及びＳ１４からＳ１７までの動作は時点ｔ２から時点Ｔ３に至るまでの制御動作である。
【０１６７】
ステップＳ２５では、第１所定時間Ｔａが経過した直後に限り、上記第１、第２トルク値Ｔｒ１，Ｔｒ２をそれぞれ前述の特性近似式ｆ１，ｆ２に代入し、パルス数Ｎ１，Ｎ２、及びその偏差Δｎを求める。次いで、ステップＳ２６で、同じく第１所定時間Ｔａが経過した直後に限り、第２タイマーｔｉｍ２をセットする。
【０１６８】
そして、ステップＳ２７で、該第２タイマｔｉｍ２が第２所定時間Ｔｂを過ぎたか否かを判定し、まだ過ぎていないとき、つまり時点ｔ２から時点ｔ３までの間は、ステップＳ２８に進んで、この第２所定時間Ｔｂの間に上記パルス数偏差Δｎだけ三層弁２２０のスリーブ２２２が移動するように、ステップモータ２５１への出力信号のパルス数偏差ΔＮを設定する。
【０１６９】
次いで、ステップＳ２９でオンオフソレノイドバルブ２１３に対する指示値をオンとし、またステップＳ３０で点火時期Ｉｇを所定値ΔＩｇだけリタードしてトルクダウンを図る。
【０１７０】
次いで、ステップＳ３１でロークラッチデューティ率ＤＬを第２デューティ率ＤＬ２とし、またステップＳ３２でハイクラッチデューティ率ＤＨとして、上記第２所定時間Ｔｂ及び次の第３所定時間の間に、該ハイクラッチデューティ率ＤＨが上記第２デューティ率ＤＨ２から第３デューティ率ＤＨ３まで変化するような値に設定する。ここで、この第３デューティ率ＤＨ３の値は１００％とされている。
【０１７１】
そして、ステップＳ１４〜Ｓ１７において、ステップモータ２５１、デューティソレノイドバルブ２７１，２７２、オノフソレノイドバルブ２１３、及びエンジン１の点火時期をそれぞれ制御することにより、前述したような時点ｔ２から所定時間Ｔｂが経過する時点ｔ３までの制御動作が得られる。
【０１７２】
一方、上記ステップＳ２７で、第２所定時間Ｔｂが経過したと判定されたときは、該ステップＳ２７からＳ３３に進み、ここで第２タイマーｔｉｍ２が上記第２所定時間Ｔｂと第３所定時間Ｔｃとの和以上に大きいか否かが判定される。そして、ＮＯのとき、つまり時点ｔ３から時点Ｔ４の間は、ステップＳ３４に進み、ここで、再びパルス数偏差ΔＮが０とされ、また、ステップＳ３５でオンオフソレノイドバルブに対する指示信号としてＯＦＦが設定される。
【０１７３】
以上の制御により、前述したような時点ｔ３から所定時間Ｔｃが経過する時点ｔ４までの制御動作が得られる。
【０１７４】
そして、第３所定時間Ｔｃも経過して時点ｔ４に至ったのちは、上記ステップＳ３３でＮＯと判定されて、該ステップＳ３３からＳ３６に進む。そして、該ステップＳ３６でハイクラッチデューティ率ＤＨが１００％の第３デューティ率ＤＨ３が代入され、またステップＳ３７でロークラッチデューティ率ＤＬが０％の第２デューティ率ＤＬ２に維持される。
【０１７５】
これにより、ハイクラッチ７０が完全締結状態となり、モードが完全にハイモードに切り換わる時点ｔ４以降の制御動作が得られ、変速比の通常のフィードバック制御が再開されて、走行状態に応じた変速比制御が実行されることになる。
【０１７６】
なお、以上においては、第１所定時間Ｔａを油温によって補正し、且つステップＳ１９でこれを第１タイマーｔｉｍ１により計時するようにしたが、これに代えて、ローモードクラッチ６０とハイモードクラッチ７０の両方が締結状態になったか否かを判定するものとして、締結側摩擦要素の油圧が所定値以上に高くなった時点をｔ２としてもよい。
【０１７７】
さらに、両クラッチ６０，７０の入、出力側回転部材６０ａ，６０ｂ；７０ａ，７０ｂ（図１参照）間の回転のスリップ量がともに所定値以下となったときに、三層弁２２０のスリーブ２２２をパルス数偏差Δｎだけ移動させるようにしてもよい。
【０１７８】
その場合は、図１及び図７に示したように、セカンダリシャフト１３の回転数を検出する出力軸回転数センサ３１０を設け、このセンサ３１０でセカンダリシャフト１３の回転数を検出すると共に、入、出力回転数センサ３０７，３０８で入、出力ディスク２１，２２の回転数を検出する。
【０１７９】
例えば、ロークラッチ６０のスリップ量ＳＬＰ（６０）は、その入力側回転部材６０ａの回転数Ｗ（６０ａ）と出力側回転部材６０ｂの回転数Ｗ（６０ｂ）とから次式のように表される。
【０１８０】
【数６】

ここで、ロークラッチ６０の入力側回転部材６０ａの回転数Ｗ（６０ａ）は、入力ディスク２１の回転数Ｗ（２１）と、ローモードギヤ列８０の第１ギヤ８１及び第２ギヤ８２の歯数Ｚ（８１），Ｚ（８２）とから次式に従って算出することができる。
【０１８１】
【数７】

また、ロークラッチ６０の出力側回転部材６０ｂの回転数Ｗ（６０ｂ）は、出力ディスク２２の回転数Ｗ（２２）と、遊星歯車機構５０のサンギヤ５２及びインターナルギヤ５３の歯数Ｚ（５２），Ｚ（５３）と、ハイモードギヤ列９０の第１ギヤ９１及び第２ギヤ９２の歯数Ｚ（９１），Ｚ（９２）と、セカンダリシャフト１３の回転数Ｗ（１３）とから次式に従って算出することができる。
【０１８２】
【数８】

各歯数Ｚは予め諸元により分かっているから、結局、ロークラッチ６０のスリップ量ＳＬＰ（６０）は、セカンダリシャフト１３の回転数Ｗ（１３）と、入、出力ディスク２１，２２の回転数Ｗ（２１），Ｗ（２２）とから算出できることになる。
【０１８３】
一方、同様に、ハイクラッチ７０のスリップ量ＳＬＰ（７０）は、その入力側回転部材７０ａの回転数Ｗ（７０ａ）と出力側回転部材７０ｂの回転数Ｗ（７０ｂ）とから次式のように表される。
【０１８４】
【数９】

ここで、ハイクラッチ７０の入力側回転部材７０ａの回転数Ｗ（７０ａ）は、出力ディスク２２の回転数Ｗ（２２）と、ハイモードギヤ列９０の第１ギヤ９１及び第２ギヤ９２の歯数Ｚ（９１），Ｚ（９２）とから次式に従って算出することができる。
【０１８５】
【数１０】

そして、ハイクラッチ７０の出力側回転部材７０ｂの回転数Ｗ（７０ｂ）は、セカンダリシャフト１３の回転数Ｗ（１３）であり、各歯数Ｚは予め諸元により分かっているから、結局、ハイクラッチ７０のスリップ量ＳＬＰ（７０）は、セカンダリシャフト１３の回転数Ｗ（１３）と、出力ディスク２２の回転数Ｗ（２２）とから算出できることになる。
【０１８６】
【発明の効果】
以上のように、本発明によれば、無段変速機構を備えたパワートレインにおいて、動力伝達経路の切換時に、その切換えに伴う変速比の変動が抑制されるから、切換前後に渡って安定して変速比を狙いの変速比に維持することができ、走行モードを切り換える際の不快なショックの発生が有効に解消される。本発明は例えばトロイダル式無段変速機構を備えたパワートレインに好ましく適用可能で、車両産業一般に広く用いられ得る。
【０１８７】
【従来の技術】
【図面の簡単な説明】
【図１】 本発明の実施の形態に係るトロイダル式無段変速機の機械的構成を示す骨子図である。
【図２】 同変速機の要部の具体的構造を展開状態で示す平面図である。
【図３】 図２のＡ−Ａ線に沿う断面図である。
【図４】 同変速機の油圧制御の回路図である。
【図５】 図３のＢ方向からみた変速制御用三層弁の周辺の部分断面図である。
【図６】 図３のＣ方向からみた変速制御機構周辺の部分断面図である。
【図７】 パワートレイン全体の制御システムを示すブロック図である。
【図８】 ステップモータのパルス数とトロイダルレシオとの関係を示す特性図である。
【図９】 ステップモーターのパルス数とユニットレシオとの関係を示す特性図である。
【図１０】 循環トルクの説明図である。
【図１１】 ステップモータのパルス数とトルクとの関係を示す特性図である。
【図１２】 ユニットレシオとトロイダルレシオとの関係を示す特性図である。
【図１３】 前進ローモードで用いられる変速線図である。
【図１４】 前進ハイモードで用いられる変速線図である。
【図１５】 後退ローモードで用いられる変速線図である。
【図１６】 正駆動状態でローモードからハイモードへの切換制御動作のタイムチャートである。
【図１７】 同制御で用いられるトルクとモード切換え開始判定用トロイダルレシオとの関係を示すマップである。
【図１８】 同制御で用いられる油温と第１所定時間との関係を示すマップである。
【図１９】 同制御で用いられるトルクとリタード量との関係を示すマップである。
【図２０】 同制御動作の一部を示すフローチャートである。
【図２１】 同じく一部を示すフローチャートである。
【図２２】 同制御動作で用いられる指数と、ステップモータへ出力する制御信号のパルス数との関係を示すマップである。
【図２３】 同制御動作の一部を示すフローチャートである。
【図２４】 同じく一部を示すフローチャートである。
【図２５】 同じく一部を示すフローチャートである。
【符号の説明】
１ エンジン
１０ パワートレイン
２０，３０ 無段変速機構
５０ 遊星歯車機構（歯車機構）
６０ ロークラッチ（摩擦要素）
７０ ハイクラッチ（摩擦要素）
２００ 油圧制御回路
２５１，２５２ ステップモータ（変速比制御手段）
２７１ ロークラッチ用デューティソレノイドバルブ（経路切換手段）
２７２ ハイクラッチ用デューティソレノイドバルブ（経路切換手段）
３００ コントロールユニット（切換制御手段）[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a vehicle power train, and more particularly to a power train control device using a continuously variable transmission mechanism, and belongs to the technical field of vehicle drive devices.
[0002]
[Prior art]
In recent years, a power train mounted on a vehicle using a continuously variable transmission mechanism has been put into practical use. Japanese Patent Laid-Open No. 9-210191 discloses a typical example. The powertrain has a toroidal-type continuously variable transmission mechanism and a planetary gear mechanism. As a power transmission path between the engine and the drive wheels, the first path that passes through both these mechanisms and the continuously variable transmission mechanism only. And a second route to Power is transmitted between the engine and the drive wheels using either one of the first and second paths by selectively fastening the friction elements for achieving each path.
[0003]
In this case, by controlling the continuously variable transmission mechanism to a predetermined speed ratio with the first path selected, a geared neutral state in which the output rotation to the drive wheel side becomes zero is obtained. By increasing or decreasing the speed ratio of the continuously variable transmission mechanism from the state, a forward state or a reverse state having a relatively low final speed ratio as a power train is realized (low mode). On the other hand, in a state where the second path is selected, a forward state with a relatively high gear ratio is realized (high mode) in which the final gear ratio as the power train changes only in accordance with the gear ratio of the continuously variable transmission mechanism.
[0004]
Generally, in a toroidal-type continuously variable transmission mechanism, the transmission ratio of the continuously variable transmission mechanism is increased in the deceleration direction in order to keep the width of tilting of the power roller interposed between the input and output disks within the smallest possible range. In the low mode, the final speed ratio of the power train is reduced in the speed increasing direction, and in the high mode, the speed ratio is increased in the decelerating direction. Therefore, some speed ratios on the deceleration side of the continuously variable transmission mechanism can achieve the same final speed ratio in both the low mode and the high mode. When the speed ratio is at this point, the low mode and the high mode Is switched, the final gear ratio does not fluctuate significantly before and after the mode switching, and the switching shock is suppressed.
[0005]
The above publication discloses that both the friction element for achieving the low mode and the friction element for achieving the high mode are both brought into the engaged state during the mode switching operation. By engaging both the friction elements, the transmission gear ratio is firmly maintained at the switching point, and the switching shock is reliably avoided.
[0006]
[Problems to be solved by the invention]
By the way, for example, in the case of the toroidal-type continuously variable transmission mechanism disclosed in the above publication, the gear ratio control is generally performed by tilting the power roller interposed between the input disk having the toroidal surface and the output disk with respect to both disks. This is done by controlling. By controlling the position of the power roller with respect to the disk, the roller inclines in an opposite direction and tilts by a predetermined angle in response to the rotation of the output disk. Position control of the roller with respect to the disk is performed by hydraulic control with respect to the support member of the roller, and the hydraulic control is executed through, for example, sleeve position control of a three-layer valve or pulse number control of a step motor that moves the sleeve. Is done.
[0007]
At this time, the gear ratio of the continuously variable transmission mechanism and the sleeve position of the three-layer valve or the pulse number of the step motor theoretically correspond one-on-one. Therefore, when the target gear ratio is set, the sleeve position of the three-layer valve, the number of pulses of the step motor, and the like for realizing it are determined in advance.
[0008]
However, deformation of the toroidal disk surface and rollers due to the torque passing through the continuously variable transmission mechanism, or dragging of the rollers due to the rotational force of the disk occurs, and in reality, a control signal with the same number of pulses is output to the step motor. However, different gear ratios may be obtained, or the number of pulses for realizing the same target gear ratio may be different. Such a deviation from the theoretical value of the gear ratio not only becomes more noticeable as the torque increases, but also the direction of deviation reverses depending on the direction of torque passage, for example, continuously variable transmission between low mode and high mode. When the transmission direction of the torque passing through the mechanism is reversed, the gear ratio is shifted in the speed increasing direction, but at the same time the mode is changed, it is changed to a shift in the deceleration direction all at once. As a result, when shifting from the low mode state to the high mode state after both the friction elements are fastened together, a significant change in the gear ratio occurs, and an unpleasant shock occurs here.
[0009]
The present invention addresses the above problems when switching between a low mode and a high mode having different power transmission paths, and reliably suppresses fluctuations in the gear ratio during the switching operation. The problem is to eliminate the problem.
[0010]
[Means for Solving the Problems]
In order to solve the above problems, the present invention uses the following means.
[0011]
  First, the invention according to claim 1 of the present application includes a first path that passes through the continuously variable transmission mechanism and the gear mechanism, and a second path that passes only through the continuously variable transmission mechanism. Route switching means for selectively switching these routes is provided.In addition, the continuously variable transmission mechanism has an input disk and an output disk, a power roller interposed between the two disks, and a trunnion that supports the power roller.A powertrain control device,The tilt angle of the power roller with respect to the input disk and the output disk is controlled by controlling the position of the trunnion with respect to the input disk and the output disk.Transmission ratio control means for controlling the transmission ratio of the continuously variable transmission mechanism;Torque detecting means for detecting torque;the aboveRouteThe speed ratio control means is configured to suppress a change in speed ratio accompanying the switching when the path is switched by the switching means.Position control amount of the above trunnion bySwitching control means for controllingThe switching control unit increases the correction amount of the trunnion position control amount to the deceleration side or the acceleration side as the torque detected by the torque detection unit increases.It is characterized by that.
[0012]
According to the present invention, when the driving mode is switched, the change in the gear ratio accompanying the switching is suppressed, so that an unpleasant shock is generated even when the mode transitions from low to high or from high to low. Is reduced.
[0014]
  Also,According to the present invention, since the gear ratio control amount is corrected according to the torque, for example, even when the torque is large and a deviation from the theoretical value of the gear ratio becomes significant, the deviation or fluctuation is not detected. Suppressed sufficiently effectively.
[0015]
  Next, the claim2The invention according to claim 1 is the invention according to claim 1,the aboveThe switching control means is configured to switch between the first path and the second path and when switching from the second path to the first path.Correction of the above-mentioned trunnion position control amountIt is characterized by reversing the direction.
  The invention according to claim 3 is the invention according to claim 1 or 2, wherein the switching control means is configured such that when the vehicle is driven by an engine in a positive drive state and when the engine is running inertia of the vehicle. In the reverse drive state driven by the above, the correction direction of the trunnion position control amount is reversed.
[0016]
  Claim 2According to the invention, for example, the transmission direction of the torque passing through the continuously variable transmission mechanism is reversed between the low mode and the high mode, and as a result, the shift direction of the speed ratio is reversed between the speed increasing side and the speed reducing side. Can handle the situation appropriately.
  According to the third aspect of the present invention, for example, the transmission direction of the torque passing through the continuously variable transmission mechanism is reversed between the forward drive state and the reverse drive state. It is possible to appropriately deal with situations where the situation is reversed with the other side.
[0017]
  Next, the invention according to claim 4 is the invention according to any one of claims 1 to 3, wherein the friction element and the second path that achieve the first path by being fastened respectively are provided. A friction element to achieve,the aboveThe path switching means is configured to perform path switching by selectively fastening one of the friction elements through a state in which both the friction elements are fastened together,the aboveThe switching control means is the aboveRouteBy switching meansAfter a predetermined time has elapsed since the start of the path switching, the position control amount of the trunnion by the speed ratio control meansIt is characterized by controlling.
[0018]
According to the present invention, since the switching of the route passes through the state in which the friction elements for achieving the first and second routes are both fastened, the speed change ratio is maintained at the switching point during that time, and the switching shock is surely achieved. To be avoided.
[0019]
Since the gear ratio control means is controlled when both the friction elements are engaged, the gear ratio does not fluctuate during the control of the gear ratio control means. The gear ratio control means can be controlled, and after that, when the mode transition starts, the target gear ratio is realized.
[0021]
  Also,Specifically, according to the present invention, it is possible to determine whether or not both friction elements are in an engaged state by measuring time.
[0022]
  Next, the claim5The invention described in claim 14In the invention described inthe aboveThe switching control means, when the fastening hydraulic pressure of the fastening side friction element becomes a predetermined value or more,Position control amount of the trunnion by the transmission ratio control meansIt is characterized by controlling.
[0023]
According to the present invention, more specifically, it can be determined whether or not both friction elements are in a state of being fastened together, depending on the progress of the fastening operation of the fastening side friction elements.
[0024]
  Next, the claim6The invention described in claim 14In the invention described inthe aboveThe switching control means, when both the engagement side friction element input, the differential rotation of the output rotation member, and the release side friction element input, the differential rotation of the output rotation member are both below a predetermined value,Position control amount of the trunnion by the transmission ratio control meansIt is characterized by controlling.
[0025]
More specifically, according to the present invention, it is possible to determine whether or not both friction elements are in a state of being fastened together with almost no slip amount of both friction elements.
[0026]
  Next, the claim7The invention according to claim 4 is provided with torque reduction means for reducing torque in the invention according to claim 4,the aboveThe switching control means is arranged so that the torque decreases when the fastening side friction element is fastened.torqueThe lowering means is controlled.
[0027]
According to this invention, the engagement shock of the engagement side friction element is reduced, and in addition, the degree of deviation from the theoretical value of the transmission ratio is suppressed.
[0030]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, the present invention will be described in more detail through embodiments of the invention.
[0031]
As shown in FIGS. 1 and 2, the power train 10 according to the present embodiment includes an input shaft 11 connected to the output shaft 2 of the engine 1 via a torsional damper 3, and an idle outside the shaft 11. It has a fitted hollow primary shaft 12 and a secondary shaft 13 arranged in parallel to these shafts 11 and 12, and these shafts 11 to 13 all extend in the lateral direction of the vehicle. Has been placed.
[0032]
In addition, on the axes of the input shaft 11 and the primary shaft 12, the toroidal first and second continuously variable transmission mechanisms 20 and 30 and a loading that enables transmission of power by applying axial loads thereto. A cam mechanism 40 is disposed, and a planetary gear mechanism 50, a low clutch 60, and a high clutch 70 are disposed on the axis of the secondary shaft 13. Further, a low mode gear train 80 and a high mode gear train 90 are interposed between the axis of the input shaft 11 and the primary shaft 12 and the axis of the secondary shaft 13.
[0033]
The first and second continuously variable transmission mechanisms 20 and 30 have substantially the same configuration, and both have input disks 21 and 31 and output disks 22 and 32 whose opposing surfaces are toroidal surfaces. Two power rollers 23 and 33 for transmitting power between the disks 21 and 22 and between the disks 32 and 32 are provided between the opposing toroidal surfaces.
[0034]
The first continuously variable transmission mechanism 20 disposed farther from the engine 1 has the input disk 21 disposed on the non-engine side, the output disk 22 disposed on the engine side, and disposed closer to the engine 1. In the second continuously variable transmission mechanism 30, the input disk 31 is disposed on the engine side and the output disk 32 is disposed on the non-engine side, and the input disks 21 and 31 of both the transmission mechanisms 20 and 30 are disposed at both ends of the primary shaft 12. The output disks 22 and 32 are integrated with each other and are rotatably supported by the intermediate portion of the primary shaft 12.
[0035]
Further, a first gear 81 constituting the low mode gear train 80 is coupled to the end of the input shaft 11 on the side opposite to the engine, and between the first gear 81 and the input disk 21 of the first continuously variable transmission mechanism 20. The loading cam mechanism 40 is interposed, and the high mode gear train 90 is formed on the outer periphery of the output disks 22 and 32 integrated with the first and second continuously variable transmission mechanisms 20 and 30. One gear 91 is provided.
[0036]
On the other hand, a second gear 82 constituting the low mode gear train 80 is rotatably supported at the end of the secondary shaft 13 on the non-engine side, and is connected to the first gear 81 via an idle gear 83. At the same time, the planetary gear mechanism 50 is disposed at an intermediate portion of the secondary shaft 13. A low clutch 60 that connects or disconnects the pinion carrier 51 of the planetary gear mechanism 50 and the second gear 82 of the low mode gear train 80 is interposed.
[0037]
Further, a second gear 92 constituting a high mode gear train 90 is rotatably supported on the engine side of the planetary gear mechanism 50, and the output disks 22, 30 in the first and second continuously variable transmission mechanisms 20, 30 are supported. The second gear 92 and the sun gear 52 of the planetary gear mechanism 50 are connected to each other, and the internal gear 53 of the planetary gear mechanism 50 is further engaged with the first gear 91 provided on the outer periphery of the planetary gear mechanism 32. It is coupled to the secondary shaft 13. A high clutch 70 that connects or disconnects the second gear 92 of the high mode gear train 90 and the secondary shaft 13 is interposed on the engine side of the planetary gear mechanism 50.
[0038]
Further, a differential device 5 is connected to an end of the secondary shaft 13 on the engine side via an output gear train 4 including first and second gears 4a and 4b and an idle gear 4c. Drive shafts 6a and 6b extending from left to right are connected to left and right drive wheels (not shown).
[0039]
An oil pump 100 is disposed at the end of the input shaft 11 on the side opposite to the engine, and is driven by the input shaft 11 via the first gear 81 of the low mode gear train 80.
[0040]
Next, the configuration of the first and second continuously variable transmission mechanisms 20 and 30 will be described in more detail using the first continuously variable transmission mechanism 20 as an example.
[0041]
As shown in FIG. 3, the pair of power rollers 23 and 23 are supported by trunnions 25 and 25 via shafts 24 and 24 extending substantially in the radial direction of the input and output disks 21 and 22, respectively. , 22 on the circumference of the toroidal surfaces opposite to each other at 180 ° opposite to each other in a substantially horizontal position and vertically in parallel, and the two disks 21 and 22 are arranged at two locations 180 ° opposite to the circumferential surface. Each toroidal surface is in contact.
[0042]
The trunnions 25 and 25 are supported between the left and right support members 26 and 26 attached to the case 101 of the power train 10 and are tangential to both the disks 21 and 22 and the shafts of the power rollers 23 and 23. Rotation about a horizontal axis X, X perpendicular to 24, 24 and linear reciprocation in the X, X direction are possible. The trunnions 25, 25 are provided with rods 27, 27 extending sideways along the axial centers X, X. The rods 27, 27 are disposed on the side surface of the case 101. A shift control unit 110 that tilts the power rollers 23 and 23 via the trunnions 25 and 25 is attached.
[0043]
The speed change control unit 110 includes a hydraulic control unit 111 and a trunnion drive unit 112, and the trunnion drive unit 112 has a speed increase attached to each of the rods 27 and 27 of the upper and lower trunnions 25 and 25 so as to face each other. The pistons 113 and 114 for use and the speed reduction are arranged, and the speed increasing and speed reducing hydraulic chambers 115 and 116 are formed by the opposed pistons 113 and 114, respectively.
[0044]
For the trunnion 25 located above, the acceleration hydraulic chamber 115 is disposed on the power roller 23 side, the deceleration hydraulic chamber 116 is disposed on the counter-power roller 23 side, and the trunnion 25 located on the lower side. The acceleration hydraulic chamber 115 is disposed on the side opposite to the power roller 23, and the deceleration hydraulic chamber 116 is disposed on the power roller 23 side.
[0045]
The speed increasing hydraulic pressure PH generated by the hydraulic pressure control unit 111 is supplied to the speed increasing hydraulic pressure chambers 115, 115 of the upper and lower trunnions 25, 25 via the oil passages 117, 118, and the same hydraulic control is performed. The deceleration hydraulic pressure PL generated by the section 111 is supplied to the deceleration hydraulic chambers 116 and 116 of the upper and lower trunnions 25 and 25 via an oil passage (not shown), and the transmission mechanism is controlled by controlling the hydraulic pressures PH and PL. The gear ratios 20 and 30 are controlled.
[0046]
Here, the specific operation of the gear ratio control for the first continuously variable transmission mechanism 20 will be described. First, the hydraulic control unit 111 shown in FIG. 3 sets the speed increasing hydraulic chambers 115, 115 of the upper and lower trunnions 25, 25. When the supplied acceleration hydraulic pressure PH is set higher than a predetermined balanced state with respect to the deceleration hydraulic pressure PL supplied to the deceleration hydraulic chambers 116, 116, the upper trunnion 25 is On the right side, the lower trunnion 25 moves horizontally to the left side.
[0047]
At this time, assuming that the illustrated output disk 22 is rotating in the c direction, the upper power roller 23 receives a downward force from the output disk 22 due to the movement to the right, and the front side of the drawing. Thus, an upward force is received from the input disk 21 rotating in the anti-c direction. The lower power roller 23 receives an upward force from the output disk 22 and receives a downward force from the input disk 21 due to the leftward movement.
[0048]
As a result, the upper and lower power rollers 23, 23 are tilted so that the contact position with the input disk 21 moves outward in the radial direction and the contact position with the output disk 22 moves inward in the radial direction. The gear ratio of 20 becomes smaller (speed increase).
[0049]
Contrary to the above, the deceleration hydraulic pressure PL supplied to the deceleration hydraulic chambers 116, 116 of the upper and lower trunnions 25, 25 is the acceleration hydraulic pressure supplied to the acceleration hydraulic chambers 115, 115. When the height of the trunnion is relatively higher than the predetermined balance state with respect to PH, the upper trunnion 25 moves horizontally on the left side and the lower trunnion 25 moves horizontally on the right side. Therefore, the lower power roller 23 receives a downward force from the output disk 22 and an upward force from the input disk 21. As a result, the upper and lower power rollers 23, 23 are tilted so that the contact position with the input disk 21 moves inward in the radial direction and the contact position with the output disk 22 moves outward in the radial direction. The gear ratio of 20 increases (deceleration).
[0050]
The supply control of the acceleration and deceleration hydraulic pressures PH and PL by the hydraulic control unit 111 will be further described in the description of the hydraulic control circuit described later.
[0051]
The configuration and operation of the first continuously variable transmission mechanism 20 as described above are the same for the second continuously variable transmission mechanism 30. As shown in FIGS. 1 and 2, the input disks 21 and 31 of the first and second continuously variable transmission mechanisms 20 and 30 are provided at both ends of the hollow primary shaft 12 loosely fitted on the input shaft 11. Are respectively spline-fitted so that these input disks 21 and 31 always rotate in the same direction, and the output disks 22 and 32 of both transmission mechanisms 20 and 30 are integrated as described above. Therefore, the rotational speeds on the output side of both transmission mechanisms 20 and 30 are always the same. Therefore, the transmission ratio control of the first and second continuously variable transmission mechanisms 20 and 30 by the hydraulic control of the power rollers 23 and 33 as described above is also performed so that the transmission ratio is always kept the same.
[0052]
Next, a hydraulic control circuit of the power train 10 constituted by the shift control unit 110 and a clutch control unit 120 (see FIG. 3) attached to the lower part of the case 101 will be described.
[0053]
As shown in FIG. 4, the hydraulic control circuit 200 includes a regulator valve 202 that adjusts the pressure of the hydraulic oil discharged from the oil pump 100 to a predetermined line pressure and outputs it to the main line 201, and the main line 201. A relief pressure 204 that generates a predetermined relief pressure using the line pressure supplied from the engine as a source pressure and outputs the relief pressure to the relief pressure line 203, and a D range, an R range, an N range, and a P range by a driver switching operation. A manual valve 205 that enables selection is provided.
[0054]
Among these valves, the manual valve 205 communicates the main line 201 with the first and second output lines 206 and 207 in the D range and the first and third output lines 206 and 208 in the R range, respectively. In the N range and P range, the line pressure is cut off.
[0055]
The regulator valve 202 and the relief valve 204 are provided with a linear solenoid valve for line pressure control 209 and a linear solenoid valve for relief pressure control 210, respectively, and a constant pressure using the discharge pressure of the pump 100 as an original pressure. The linear solenoid valves 209 and 210 generate control pressures based on a constant pressure generated by the reducing valve 211, respectively.
[0056]
These control pressures are supplied to the control ports 202a and 204a of the regulator valve 202 and the relief valve 204, so that the line pressure and the relief pressure are respectively controlled by control signals output to the linear solenoid valves 209 and 210, respectively. Will be adjusted.
[0057]
Further, the constant pressure generated by the reducing valve 211 is also led to an on / off solenoid valve 213 that operates the fail-safe valve 212. The on / off solenoid valve 213 is normally turned on to supply the constant pressure to the control port 212a of the fail safe valve 212, thereby moving the spool of the valve 212 to the right side, while at the time of fail safe, etc. When the valve is turned off, the constant pressure is off-drained from the control port 212a of the fail-safe valve 212, thereby moving the spool of the valve 212 to the left.
[0058]
In addition, the hydraulic pressure control circuit 200 includes three forward speed generating pressures PH and deceleration hydraulic pressures PL for shifting control based on the line pressure and the relief pressure, respectively, at the time of forward movement and at the time of backward movement. A layer valve 220 and a reverse three-layer valve 230 and a shift valve 240 for selectively operating these three-layer valves 220 and 230 are provided.
[0059]
The position of the spool of the shift valve 240 is determined by whether or not the line pressure is supplied as a control pressure to the control port 240a at one end. When the line pressure is not supplied, the spool is positioned on the right side. The main line 201 is communicated with a line pressure supply line 241 that communicates with the forward three-layer valve 220. When the line pressure is supplied, the spool is positioned on the left side, and the main line 201 is connected with the reverse three-layer valve. It operates to communicate with a line pressure supply line 242 leading to 230.
[0060]
Here, the line pressure is supplied to the control port 240a of the shift valve 240 under normal conditions when the manual valve 205 is connected via the fail-safe valve 212 and the third output line 208 whose spool has moved to the right side. When it is located in the R range. On the other hand, even when the spool of the fail safe valve 212 is moved to the right side at the normal time, the line pressure is not supplied to the control port 240a of the shift valve 240 when the manual valve 205 is positioned in the D range. Further, at the time of fail safe, the spool of the fail safe valve 212 moves to the left side, and the shift valve 240 and the third output line 208 are shut off, so that even if the manual valve 205 is located in the R range, the shift valve 240 Line pressure is not supplied to the control port 240a.
[0061]
The three-layer valves 220 and 230 for forward movement and backward movement have the same configuration, and the sleeves 222 and 232 are fitted to the bores 221 and 231 so as to be movable in the axial direction. The spools 223 and 233 are respectively fitted so as to be movable in the direction, and both are accommodated in the valve body 111a of the hydraulic control unit 111 in the transmission control unit 110 shown in FIG.
[0062]
In addition, line pressure ports 224 and 234 to which line pressure supply lines 241 and 242 led from the shift valve 240 are connected are provided at the center of these three-layer valves 220 and 230, and both end portions are provided. Are provided with first and second relief pressure ports 225, 226, 235 and 236 to which the relief pressure line 203 is branched and connected respectively. Further, between the line pressure ports 224 and 234 and the first relief pressure ports 225 and 235, speed increasing pressure ports 227 and 237 are provided between the line pressure ports 224 and 234 and the second relief pressure ports 226 and 236, respectively. Between them, deceleration pressure ports 228 and 238 are respectively provided.
[0063]
The lines 243 and 244 led from the acceleration pressure ports 227 and 237 of the forward and backward three-layer valves 220 and 230, respectively, and the deceleration pressure ports 228 and 228 of the forward and backward three-layer valves 220 and 230, respectively. Lines 245 and 246 led from 238 are connected to the shift valve 240, and when the spool of the shift valve 240 is located on the right side, the acceleration pressure port 227 and the deceleration of the forward three-layer valve 220 are reduced. Lines 243 and 245 led from the pressure port 228 are connected to the acceleration line 247 and the deceleration line 248, respectively, and communicate with the acceleration hydraulic chambers 115 and 115 and the deceleration hydraulic chambers 116 and 116, respectively.
[0064]
Further, when the spool of the shift valve 240 is positioned on the left side, the lines 244 and 246 led from the acceleration pressure port 237 and the deceleration pressure port 238 of the reverse three-layer valve 230 are the acceleration line 247 and the deceleration line. The speed increasing hydraulic chambers 115 and 115 and the speed reducing hydraulic chambers 116 and 116 communicate with the line 248, respectively.
[0065]
Here, the operation of these three-layer valves 220 and 230 will be described with reference to FIG. In FIG. 5, the directions of the three-layer valves 220 and 230 are opposite to those in FIG. 4. As shown in the drawing, when the sleeve 222 of the forward three-layer valve 220 moves relatively to the left side (in the direction of the arrow g), for example, from the state where the positional relationship between the sleeve 222 and the spool 223 is in the neutral position, the line pressure The degree of communication between the port 224 and the acceleration pressure port 227 and the degree of communication between the second relief pressure port 226 and the deceleration pressure port 228 are increased, and the sleeve 222 is relatively moved to the right (arrow h direction). Then, the degree of communication between the line pressure port 224 and the deceleration pressure port 228 and the degree of communication between the first relief pressure port 225 and the acceleration pressure port 227 are increased. Therefore, in the former case, the acceleration hydraulic pressure PH increases and the deceleration hydraulic pressure PL decreases, and in the latter case, the deceleration hydraulic pressure PL increases and the acceleration hydraulic pressure PH decreases.
[0066]
The above operation is the same for the reverse three-layer valve 230, and step motors 251 and 252 for operating the sleeves 222 and 232 of the three-layer valves 220 and 230 as described above are provided, respectively. It is connected to sleeves 222 and 232 of the forward and backward three-layer valves 220 and 230 through 253 and 254.
[0067]
In addition, a cam mechanism 260 that moves the spools 223 and 233 in the axial direction against the spring force of the springs 229 and 239 in accordance with the movement of the sleeves 222 and 232 by the step motors 251 and 252 is provided.
[0068]
As shown in FIGS. 5 and 6, the cam mechanism 260 has one end face formed as a spiral cam face 261 a and the end of the rod 37 of the trunnion 35 positioned above the second continuously variable transmission mechanism 30. Precise cam 261 attached to the part and one end side of spools 223, 233 of forward and backward three-layer valves 220, 230 are arranged in a direction orthogonal to these, and rotate to valve body 111a of hydraulic control part 111 A freely supported shaft 262, a driven lever 263 attached to one end portion of the shaft 262 and having a swing end abutting against the cam surface 261a of the recess cam 261, and a shaft 262 are also attached to The moving end engages with the notches 223a and 233a provided at one end of the spools 223 and 233 of the forward and backward three-layer valves 220 and 230. It is composed of a forward and driving lever 264, 265 for retracting that.
[0069]
Then, when the upper power roller 33 in the second continuously variable transmission mechanism 30 is tilted by the control of the acceleration hydraulic pressure PH and the deceleration hydraulic pressure PL, the trunnion 35 and the rod 37 positioned above are accordingly moved. By rotating integrally around the shaft center X, the recess cam 261 also rotates integrally therewith, and the driven lever 263 whose swing end abuts on the cam surface 261a swings a predetermined amount, and the shaft The forward and backward drive levers 264 and 265 are also oscillated by the same angle via the 262, and as a result, the spools 223 of the forward and backward three-layer valves 220 and 230 by an amount corresponding to the oscillating angle. 233 moves in the axial direction.
[0070]
Therefore, the positions of these spools 223, 233 are the tilt angles of the power roller 33 of the second continuously variable transmission mechanism 30 (and the power roller 23 of the first continuously variable transmission mechanism 20), in other words, the transmission mechanisms 20 thereof. , 30 at all times.
[0071]
Here, the control operation of the transmission ratio (toroidal ratio Rt) of the continuously variable transmission mechanisms 20 and 30 will be described taking forward travel as an example.
[0072]
First, the control pressures of the regulator valve 202 and the relief valve 204 are generated by the line pressure control linear solenoid valve 209 and the relief pressure control linear solenoid valve 210 in the hydraulic control circuit 200, and the line pressure corresponding to the control pressure is determined. A relief pressure is generated.
[0073]
Of these hydraulic pressures, the line pressure is supplied from the main line 201 to the line pressure port 224 of the three-layer valve 220 via the shift valve 240 and the line 241. The relief pressure is supplied to the first and second relief pressure ports 225 and 226 of the three-layer valve 220 via the line 203. Based on the line pressure and the relief pressure, the step-up motor 251 controls the three-layer valve 220 to increase the speed supplied to the speed increasing hydraulic chambers 115 and 115 and the speed reducing hydraulic chambers 116 and 116 of the speed change control unit 110, respectively. The differential pressure ΔP (= PH−PL) between the speed hydraulic pressure PH and the deceleration hydraulic pressure PL is controlled.
[0074]
As shown in FIG. 6, this differential pressure control is applied to the trunnions 25, 35 of the continuously variable transmission mechanisms 20, 30 or the power rollers 23, 33 against the traction force T 1 or T 2 acting on the trunnions 25, 35. The rollers 23, 33 are held at a predetermined neutral position where the tilt does not proceed without receiving the rotation of the disks 21, 22, 31, 32, and the trunnions 25, 35 to the power rollers 23, 33 are moved from the neutral position. In this control, the rollers 23 and 33 are moved along the axial centers X and X so that the tilting progresses (the toroidal ratio changes) in response to the rotation of the disks 21, 22, 31 and 32.
[0075]
Now, for example, when the torque passes through the speed change mechanism 20, 30 from the input disk 21, 31 side to the output disk 22, 32 side, the power rollers 23, 33 are rotated by the rotation of the input disks 21, 31 in the a and a directions. Since the power rollers 23 and 33 and the trunnions 25 and 35 supporting the power rollers 23 and 33 and the trunnions 25 and 35 that support the power rollers 23 and 33 are driven in the b and b directions, the force for dragging them in the same direction as the rotation directions a and a of the input disks 21 and 31 Act. Further, since the output disks 22 and 32 are driven in the c and c directions by the rotation of the power rollers 23 and 33 in the b and b directions, the reaction force is opposite to the rotational directions c and c of the output disks 22 and 32. Directional force acts on the power rollers 23 and 33 or trunnions 25 and 35. As a result, traction forces T1 and T1 in the direction approaching the trunnion driving unit 112 act on the power rollers 23 and 33 and the trunnions 25 and 35.
[0076]
Conversely, for example, when the torque passes through the speed change mechanism 20, 30 from the output disk 22, 32 side to the input disk 21, 31 side, the power rollers 23, 33 are rotated by the rotation of the output disk 22, 32 in the c, c direction. Are driven in the b and b directions, the power rollers 23 and 33 and the trunnions 25 and 35 supporting the power rollers 23 and 33 try to drag them in the same direction as the rotation directions c and c of the output disks 22 and 32. Force acts. Further, since the input disks 21 and 31 are driven in the a and a directions by the rotation of the power rollers 23 and 33 in the b and b directions, the reaction force is opposite to the rotational directions a and a of the input disks 21 and 31. Directional force acts on the power rollers 23 and 33 or trunnions 25 and 35. As a result, the traction forces T2 and T2 in the direction away from the trunnion driving unit 112 act on the power rollers 23 and 33 and the trunnions 25 and 35.
[0077]
Therefore, in order to hold the power rollers 23 and 33 in the neutral position against the traction force T1 or T2, a differential pressure is applied to the acceleration and deceleration hydraulic chambers 115 and 116 provided in the trunnions 25 and 35, respectively. The speed increasing hydraulic pressure PH and the speed reducing hydraulic pressure PL are supplied so that ΔP has a magnitude that balances the traction force T.
[0078]
Now, for example, the toroidal ratio is reduced (increased) from this neutral state, and the sleeve 222 of the forward three-layer valve 220 is moved to the left (g direction) in FIGS. 5 and 6 by the step motor 251. Then, the degree of communication between the line pressure port 224 and the acceleration pressure port 227 of the three-layer valve 220 and the degree of communication between the second relief pressure port 226 and the deceleration pressure port 228 are increased. The acceleration hydraulic pressure PH supplied from the acceleration pressure line 247 to the acceleration hydraulic chambers 115 and 115 is increased, and the deceleration supplied from the deceleration pressure line 248 to the deceleration hydraulic chambers 116 and 116 is increased. The hydraulic pressure PL is reduced and the differential pressure ΔP increases, and as a result, the trunnions 25 and 35 or the power rollers 23 and 33 move in the directions d1 and d1 shown in FIG. It becomes door.
[0079]
As a result of this movement, the power rollers 23 and 33 are tilted so that the contact positions with the input disks 21 and 31 are displaced outward in the radial direction and the contact positions with the output disks 22 and 32 are displaced inward in the radial direction. In turn, the first and second continuously variable transmission mechanisms 20, 30 are increased in speed, and the toroidal ratio is reduced.
[0080]
Further, the tilt of the power roller 33 in the second continuously variable transmission mechanism 30 as described above causes the recess cam 261 in the cam mechanism 260 to rotate by the same angle in the same direction (direction e shown in FIG. 5). The driven lever 263, the shaft 262, and the drive lever 264 in the cam mechanism 260 all rotate in the direction f shown in FIG.
[0081]
As a result, the spool 223 of the three-layer valve 220 is moved in the g direction by the spring force of the spring 229, that is, in the left direction in FIGS. 5 and 6. In this direction, the sleeve 222 is moved by the step motor 251. Therefore, as described above, the increased degree of communication between the line pressure port 224 and the speed increasing pressure port 227 and the degree of communication between the second relief pressure port 226 and the deceleration pressure port 228 are initially set as described above. Will return to the neutral state.
[0082]
As a result, the differential pressure ΔP becomes smaller again, and the speed change operation as described above is completed. The speed change ratio of the continuously variable transmission mechanisms 20 and 30, that is, the toroidal ratio is changed by a predetermined amount, and the rollers 23 and 33 are neutral. It will return to the position again and be held.
[0083]
In this case, the speed change operation ends when the spool 223 moves to a predetermined neutral position in the positional relationship with the sleeve 222. The position is moved by the step motor 251 through the sleeve 222. The position of the sleeve 222 corresponds to the tilt angle of the power rollers 23 and 33 and the trunnions 25 and 35 via the cam mechanism 260. This corresponds to tilt angles of 25 and 35. As a result, the control amount of the step motor 251 corresponds to the gear ratio of the first and second continuously variable transmission mechanisms 20 and 30, and the toroidal ratio is controlled by the pulse control for the step motor 251. .
[0084]
The above operation is similarly performed when the sleeve 222 of the three-layer valve 220 is moved to the right side (h direction) in the opposite direction in FIGS. 5 and 6 by the step motor 251, and in this case, the trunnions 25, 35 are used. Or the power rollers 23 and 33 move in the d2 and d2 directions shown in FIG. 6, and the toroidal ratio increases (deceleration).
[0085]
On the other hand, as shown in FIG. 4, the hydraulic control circuit 200 includes two duty solenoid valves 271 for controlling the low clutch 60 and the high clutch 70 in addition to the above-described configuration for controlling the gear ratio. 272, and the first output line 206 led from the manual valve 205 is connected to the low clutch duty solenoid valve 271 and the second output line 207 is connected to the high clutch duty solenoid valve 272, respectively. .
[0086]
Then, the line pressure from the first output line 206 is adjusted by the low clutch duty solenoid valve 271 to generate the engagement pressure (low clutch pressure) of the low clutch 60. When this is normal, the fail safe valve 212 and By being supplied to the hydraulic chamber of the low clutch 60 via the low clutch line 274, the low clutch 60 is fastened with a fastening force corresponding to the size. Also, the operation of the high clutch duty solenoid valve 272 adjusts the line pressure from the second output line 207 to generate the engagement pressure (high clutch pressure) of the high clutch 70, which is generated via the high clutch line 275. By being supplied to the hydraulic chamber of the high clutch 70, the high clutch 70 is fastened with a fastening force corresponding to its size.
[0087]
In this case, these duty solenoid valves 271 and 272 do not output the clutch pressure when the duty ratio of the control signal is 0% (fully closed), and the line pressure supplied when 100% is the clutch pressure as it is. Is output (fully open). At an intermediate duty ratio, a clutch pressure corresponding to the value is generated.
[0088]
Here, the low clutch line 274 and the high clutch line 275 are provided with accumulators 276 and 277, respectively, and by gradually supplying the fastening pressure to the low clutch 60 and the high clutch 70, the clutch 60, The occurrence of shock at the time of fastening 70 is suppressed.
[0089]
Further, as described above, the third output line 208 led from the manual valve 205 is connected to the control port 240a of the shift valve 240 via the fail safe valve 212 in the normal state, and the manual valve 205 is connected to the R range. When the position is moved to the position, the line pressure is supplied to the control port 240a of the shift valve 240, and the spool of the shift valve 240 is moved to the left side, that is, the retreat position.
[0090]
Further, at the time of fail safe or the like, the on / off solenoid valve 213 for operating the fail safe valve 212 is turned off, and the spool of the fail safe valve 212 is moved to the left side, whereby the low clutch duty solenoid valve is operated. 271 and the low clutch line 274 and between the third output line 208 and the shift valve 240 are cut off. At this time, in particular, the low clutch line 274 communicating with the hydraulic chamber of the low clutch 60 communicates with the drain port 212b of the fail safe valve 212, and the low clutch pressure is quickly discharged from the drain port 212b.
[0091]
In addition to the above configuration, the hydraulic control circuit 200 shown in FIG. 4 is provided with a lubrication line 281 led from the drain port of the regulator valve 202. The lubrication hydraulic pressure is supplied to the lubrication line 281 at a predetermined value. The relief valve 282 to be adjusted to the first and second opening / closing valves 283, 284, etc. are arranged, and the lubricating oil for each part of the power train such as the first and second continuously variable transmission mechanisms 20, 30 and the planetary gear mechanism 50 is disposed. The supply is controlled.
[0092]
The power train 10 according to the present embodiment has the above-described mechanical configuration and the hydraulic control circuit 200, and the hydraulic control circuit 200 is used to change the speed of the first and second continuously variable transmission mechanisms 20, 30. A control unit that controls the transmission ratio (unit ratio Ru) of the power train 10 as a whole by performing ratio control and engagement control of the clutches 60 and 70 is provided.
[0093]
As shown in FIG. 7, the control unit 300 includes a vehicle speed sensor 301 that detects the vehicle speed of the vehicle, an engine speed sensor 302 that detects the engine speed, a throttle opening sensor 303 that detects the throttle opening, A selection range sensor 304 for detecting a range selected by a person, an idle switch 305 for detecting non-depression of an accelerator pedal, an oil temperature sensor 306 for detecting a temperature of hydraulic oil, and input disks 21, 31 and an output disk 22, While the signals from the input rotational speed sensor 307 and the output rotational speed sensor 308 (see FIG. 1) for detecting the rotational speed of 32 are input, the running state of the vehicle indicated by the signals from these sensors and switches 301 to 308 Or for line pressure control and relief depending on engine operating conditions Linear solenoid valves 209 and 210 for control, on / off solenoid valves 213, duty solenoid valves 271 and 272 for the low clutch 60 and the high clutch 70, step motors for the forward three-layer valve 220 and the reverse three-layer valve 230 251 and 252, and the control signal is output to the spark plug 309 and the like of the engine 1.
[0094]
Next, features of the present invention will be described.
[0095]
In the power train 10 in which the continuously variable transmission mechanisms 20 and 30 are mounted, when the N range is selected, both the low clutch 60 and the high clutch 70 are released. Therefore, the power transmitted from the input shaft 11 side to the secondary shaft 13 side is not transmitted to the planetary gear mechanism 50 or the secondary shaft 13, and therefore power is not output from the differential device 5 to the drive wheels. Absent.
[0096]
At this time, in the planetary gear mechanism 50, the sun gear 52 is driven by the power from the high mode gear train 90, but the power from the low mode gear train 80 is transmitted to the rotating member 60a on the input side of the low clutch 60 (see FIG. 1). The pinion carrier 51 is not transmitted to the pinion carrier 51, and the internal gear 53 coupled to the secondary shaft 13 is fixed. Therefore, the pinion carrier 51 is not linked with the rotation of the sun gear 52. Rotating under load.
[0097]
In this state, by setting the toroidal ratio to a predetermined value, the rotational speed of the pinion carrier 51 is made equal to the rotational speed of the input and output side rotating members 60a and 60b (see FIG. 1) of the low clutch 60. Can be controlled to speed. In other words, by controlling the toroidal ratio to the predetermined value, even if the low clutch 60 is connected, the rotation of the internal gear 53 or the secondary shaft 13 can be made zero. Thereby, a so-called geared neutral (GN) state is obtained.
[0098]
Here, the relationship between the pulse number (N) of the control signal output to the stepping motors 251 and 252 and the toroidal ratio (Rt) has a characteristic as shown in FIG. 8, for example.
[0099]
That is, when the number of pulses increases (changes to the plus side), the toroidal ratio decreases (changes to the acceleration side). At this time, the sleeves 222 and 232 of the three-layer valves 220 and 230 move in the direction of the arrow g shown in FIGS. 5 and 6 as described above. At this time, the movement of the sleeves 222 and 232 in the direction g away from the pulse motors 251 and 252 is defined as the plus side.
[0100]
Conversely, when the number of pulses decreases (changes to the minus side), the toroidal ratio increases (changes to the deceleration side). At this time, the sleeves 222 and 232 of the three-layer valves 220 and 230 move in the arrow h direction shown in FIGS. 5 and 6 as described above. In this case, the movement of the sleeves 222 and 232 in the direction h approaching the pulse motors 251 and 252 is defined as the minus side.
[0101]
The toroidal ratio (GN ratio Rtn) when the geared neutral state is obtained is smaller than 1, and the number of pulses (GN pulse number Nn) for realizing the GN ratio Rtn is relatively positive.
[0102]
On the other hand, the relationship between the pulse number (N) of the control signal output to the step motors 251 and 252 and the unit ratio (Ru) has a characteristic as shown in FIG.
[0103]
That is, when the number of pulses is equal to the number of GN pulses (Nn), the unit ratio becomes infinite as indicated by the symbol a or i. When the pulse number decreases (changes to the minus side) from the GN pulse number (Nn) and the toroidal ratio increases (changes to the deceleration side), the input rotational speed to the sun gear 52 decreases, thereby causing the planetary gear mechanism. The 50 internal gears 53 begin to rotate in the forward direction. That is, as the number of pulses decreases, the sleeve position moves to the minus side, and the toroidal ratio increases, the unit ratio decreases (changes to the speed increasing side), and the forward low mode characteristic LF is realized. To do.
[0104]
Conversely, when the pulse number increases from the GN pulse number (Nn) (changes to the plus side) and the toroidal ratio decreases (changes to the acceleration side), the input rotational speed to the sun gear 52 increases, The internal gear 53 of the planetary gear mechanism 50 starts to rotate in the reverse direction. That is, as the number of pulses increases, the sleeve position moves to the plus side, and the toroidal ratio decreases, the unit ratio increases (changes to the deceleration side). LR is realized.
[0105]
In addition, after starting in the forward low mode LF, the number of pulses decreases and the unit ratio decreases, and as shown by the symbol C in FIGS. 8 and 9, a predetermined switching point (number of pulses Nm, When the toroidal ratio Rtm and the unit ratio Rum) are reached, the low clutch 60 is disengaged and the high clutch 70 is engaged. That is, the clutches 60 and 70 are switched. As a result, the power from the input shaft 11 is transmitted to the secondary shaft 13 via the first and second continuously variable transmission mechanisms 20 and 30, the high mode gear train 90 and the high clutch 70. In this state, if the gear ratio of the high mode gear train 90 is 1, the unit ratio is equal to the toroidal ratio and is substantially the same as the characteristics of the toroidal ratio shown in FIG. That is, as the number of pulses increases, the sleeve position moves to the plus side, and the toroidal ratio decreases, the unit ratio becomes smaller (changes to the higher speed side), and the forward high mode characteristic HF is realized. To do.
[0106]
Here, as shown in FIG. 10, in the forward low mode LF, the circulating torque is generated when the vehicle is driven by the engine 1 in the positive drive state. That is, as indicated by an arrow i, torque from the engine 1 is transmitted from the end on the opposite side of the input shaft 11 to the secondary shaft 13 side via the low mode gear train 80, while the planet on the secondary shaft 13 is transmitted. Torque as a reaction force generated in the gear mechanism 50 is returned to the output disks 22 and 32 of the continuously variable transmission mechanisms 20 and 30 through the high mode gear train 90 as indicated by an arrow j. The return torque is transmitted again to the low mode gear train 80 side via the input disks 21 and 31, the primary shaft 12, the loading cam mechanism 40, and the like, as indicated by an arrow k. Therefore, in the forward low mode LF, in the transmission mechanisms 20 and 30, the torque is transmitted from the output disks 22 and 32 side to the input disks 21 and 31 side in the forward drive state, and the power rollers 23 and 33. The trunnions 25 and 35 are subjected to traction forces T2 and T2 in the speed increasing direction shown in FIG. 6 for reducing the toroidal ratio.
[0107]
On the other hand, when the forward high mode HF is in the positive drive state, the rotation from the engine 1 input to the input shaft 11 is input from the loading cam mechanism 40 to the input disks 21 and 31 of the continuously variable transmission mechanisms 20 and 30. The power is transmitted to the output disks 22 and 32 via the power rollers 23 and 33, respectively, and further transmitted from the high mode gear train 90 to the secondary shaft 13 via the high clutch 70. Therefore, in the high mode HF, in the speed change mechanisms 20 and 30, torque is transmitted from the input disks 21 and 31 side to the output disks 22 and 32 side in the normal drive state, and the power rollers 23, 33 and The trunnions 25 and 35 are subjected to traction forces T1 and T1 in the deceleration direction shown in FIG. 6 for increasing the toroidal ratio.
[0108]
In the forward low mode LF, when the engine 1 is in the reverse drive state driven by the running inertia of the vehicle, the torque is applied from the input disks 21 and 31 side to the output disk 22, contrary to the forward drive state. , 32, the traction forces T1 and T1 in the deceleration direction act on the power rollers 23 and 33 and the trunnions 25 and 35, and when the forward high mode HF is in the reverse drive state, it is in the forward drive state. On the contrary, torque is transmitted from the output disks 22 and 32 side to the input disks 21 and 31 side, and the traction forces T2 and T2 in the speed increasing direction act on the power rollers 23 and 33 and trunnions 25 and 35. .
[0109]
Therefore, for example, when switching between low and high modes occurs in the same driving state of either forward or reverse, or switching between forward and reverse driving states while maintaining the same mode of either low or high. Occurs, the direction of transmission of the torque passing through the speed change mechanisms 20 and 30 is reversed, and accordingly, the direction of the traction force acting on the power rollers 23 and 33 and the trunnions 25 and 35 is also reversed.
[0110]
At this time, as described with reference to FIG. 6, the trunnions 25 and 35 of the continuously variable transmission mechanisms 20 and 30 are supplied with the differential pressure ΔP that can counteract the traction force T1 or T2, and are neutral. Although it is held in position, in reality, it is only the pistons 113 and 114 disposed in the trunnion drive unit 112 and the vicinity thereof that directly receive such differential pressure ΔP. For example, rods 27 and 37, trunnion bodies 25 and 35 arranged at positions close to input / output disks 21, 22, 31, and 32, or their connected portions, or roller support shafts 24 and 24, and The connecting portion between the shafts 24 and 24 and the trunnions 25 and 35 is bent or distorted so as to be dragged by the traction force T1 or T2, and is physically As a result, the power rollers 23 and 33 are slightly offset from the neutral position in the speed increasing direction (d1 direction: traction force T2) or the deceleration direction (d2 direction: traction force T1). is there.
[0111]
Therefore, for example, as a specific example, when switching from the low mode LF to the high mode HF occurs in the positive drive state, the power rollers 23 and 33 are offset in the acceleration direction by the traction forces T2 and T2. The position is instantaneously moved from the position to the position offset in the deceleration direction by the traction forces T1 and T1, with the original neutral position interposed therebetween.
[0112]
FIG. 11 is a graph showing the experimental results of examining how the number of pulses that can realize the mode switching point toroidal ratio Rtm or the sleeve position changes depending on the transmission torque (input torque) passing through the transmission mechanisms 20 and 30. .
[0113]
As is clear from FIG. 11, when the torque is zero, the mode as the theoretical value described above is used in the low mode LF, the high mode HF, the forward drive, and the reverse drive. The mode switching point toroidal ratio Rtm is realized at the switching point pulse number Nm or the mode switching point sleeve position Sm.
[0114]
However, as the torque increases, the number of pulses or the sleeve position capable of realizing the mode switching point toroidal ratio Rtm greatly deviates from the theoretical values Nm and Sm. For example, in the low mode LF forward drive state (f1 characteristic line in the figure), as the torque increases, the number of pulses or the sleeve position must be gradually shifted from the theoretical values Nm and Sm in the deceleration direction. The switching point toroidal ratio Rtm is not realized. This indicates that the power rollers 23 and 33 are offset in the speed increasing direction as described above in the positive drive state of the low mode LF, and the offset amount changes according to the torque. . Such characteristics are the same in the other three states (characteristic lines f2, f3, and f4 in the figure).
[0115]
Therefore, for example, when the mode is switched, if the torque in the positive drive state of the low mode LF is Tr1 and the torque in the positive drive state of the high mode HF is Tr2, the deviation of the number of pulses Δn (N2 If -N1) is not compensated, significant toroidal ratio or unit ratio fluctuation occurs simultaneously with the mode switching, and an unpleasant shock occurs.
[0116]
In the power train 10 equipped with the continuously variable transmission mechanisms 20 and 30 according to this embodiment, measures are taken to eliminate such problems.
[0117]
The above-described four characteristic lines f1 to f4 are given the following approximate expressions as functions of torque, respectively. X is a torque, and K11 to K14, K21 to K24, K31, K41, α, and β are constants.
[0118]
[Expression 1]

[0119]
[Expression 2]

[0120]
[Equation 3]

[0121]
[Expression 4]

Next, as shown in FIG. 12, the mode switching point (c) is the only point at which the same unit ratio (Rum) can be obtained with the same toroidal ratio (Rtm) in both the low mode LF and the high mode HF. Therefore, by switching the mode at this point (c), smooth mode switching without a significant change in the unit ratio before and after the switching and thus without the occurrence of switching shock is realized. Various physical quantities such as the number of switching point pulses (Nm: when torque is zero) or the switching point sleeve position (Sm: when torque is zero) for realizing this switching point toroidal ratio (Rtm) are shown in FIGS. As shown in FIG. 11, only one point is theoretically determined. As described above, the number of switching point pulses or the switching point sleeve position for realizing the switching point toroidal ratio (Rtm) changes according to the torque.
[0122]
Therefore, in theory, when the actual toroidal ratio reaches the switching point toroidal ratio (Rtm), and therefore the number of pulses or sleeve position reaches the number of switching point pulses (Nm) or the switching point sleeve position (Sm). If the control of the number of pulses or the sleeve position is stopped, the actual toroidal ratio is stabilized at the switching point toroidal ratio (Rtm). However, in the vicinity of the mode switching point (c), the switching The power rollers 23 and 33 are continuously tilted so as to approach the point. Therefore, in reality, as a result of the inertia due to the continuous tilting motion, the actual toroidal ratio is changed to the switching point toroidal ratio ( After reaching Rtm), control of the number of pulses or sleeve position was stopped. Then, the power rollers 23 and 33 are tilted too much, and as a result, the actual toroidal ratio exceeds the switching point toroidal ratio (Rtm) and is further decelerated (from the low mode LF to the high mode HF and from the high mode HF to the low mode LF). In any mode switching, the toroidal ratio changes too much from the speed increasing side to the speed reducing side before the switching (enters a region indicated by a chain line in FIG. 12).
[0123]
Since the inertia of the tilting motion of the rollers 23 and 33 increases as the transmission torque increases, the overshoot amount of the rollers 23 and 33 increases in accordance with the transmission torque. Therefore, at the time of mode switching, particularly at the start thereof, if the shift from the switching point of the actual ratio to the deceleration side due to overshooting of the power rollers 23 and 33 is not taken into account, a significant toroidal will occur as the mode switching operation proceeds. Variations in the ratio or unit ratio will occur and an unpleasant shock will occur.
[0124]
In the power train 10 on which the continuously variable transmission mechanisms 20 and 30 according to this embodiment are mounted, measures are also taken to eliminate such problems.
[0125]
Next, a specific control operation at the time of switching between the forward low mode LF and the forward high mode HF will be described in detail. This control is basically control in the vicinity of the mode switching point (c), and the low clutch 60 and the high clutch 70 serving as friction elements for achieving each mode are switched. That is, when switching from the low mode LF to the high mode HF, the low clutch 60 is disconnected and the high clutch 70 is engaged. Conversely, when switching from the high mode HF to the low mode LF, the high clutch 70 is disconnected and the low clutch 60 is engaged. The changing operation of the clutches 60 and 70 is started when the actual toroidal ratio reaches the switching point toroidal ratio (Rtm). During the changing operation, the actual toroidal ratio is constant at the switching point toroidal ratio (Rtm). The number of pulses for the forward stepping motor 251 is controlled so as to be held by the motor.
[0126]
On the other hand, during normal times other than such mode switching, basically, as shown in FIGS. 13 to 15, in each of the forward low mode LF, forward high mode HF, and reverse low mode LR, the vehicle speed ( Feedback control of the speed ratio (toroidal ratio and unit ratio) is performed based on a preset speed diagram with the vehicle running state such as V) and throttle opening (TVO) as parameters.
[0127]
In this gear ratio control, as shown in FIG. 14, first, the actual vehicle speed (V) and the actual throttle opening (TVO) are applied to the above shift diagram to obtain the target engine speed (Neo), and then After calculating the target unit ratio (Ruo) from the target engine speed (Neo) and the actual vehicle speed (V), and further setting the target toroidal ratio (Rto) to obtain this target unit ratio (Ruo), this target The toroidal ratio is feedback-controlled through pulse control (sleeve position control) for the step motors 251 and 252 so that the toroidal ratio (Rto) is realized.
[0128]
In each shift diagram, a mode switching line (M) having a slope of a mode switching point unit ratio (Rum) is shown.
[0129]
Further description will be given with reference to the time chart of FIG. As an example, this time chart shows a case where, for example, the accelerator pedal is continuously depressed and the vehicle speed increases, and as a result, switching from the low mode LF to the high mode HF occurs in the normal drive state. .
[0130]
First, during the period up to the time point t1 and during the period after the time point t4, the normal speed ratio feedback (F / B) control based on the above-described traveling state and shift diagram is performed, and the time points between these Mode switching control is executed during the period from t1 to time t4.
[0131]
First, during the period up to the time point t1, the low mode LF is achieved. Therefore, the duty ratio DL for the low clutch duty solenoid valve 271 is set to 100% (DL1) with the on-off solenoid valve 213 turned on. The duty ratio DH for the high clutch duty solenoid valve 272 is 0% (DH1). As a result, the line pressure is supplied as it is to the hydraulic chamber of the low clutch 60 as the low clutch pressure (EL), and the low clutch 60 is fully engaged, while the hydraulic chamber of the high clutch 70 has a high clutch. The pressure (EH) is not supplied, and the high clutch 70 is completely released.
[0132]
During the period up to the time point t1, the target unit ratio (Ruo) is set so as to change to the speed increasing side by the feedback control of the transmission ratio. Therefore, as indicated by the chain line in the figure, the target toroidal ratio is set. (Rto) is set to change to the deceleration side, and the toroidal ratio (Rt) is feedback-controlled so that this target toroidal ratio (Rto) is realized. As a result, as shown by the solid line in the figure, The toroidal ratio (Rtr) changes to the deceleration side so as to follow the target toroidal ratio (Rto). At this time, the number of pulses (N) and the sleeve position (S) for the three-layer valve 220 both change to the negative side.
[0133]
Then, as indicated by the reference symbol, at the time t1 when the actual toroidal ratio (Rtr) reaches a predetermined toroidal ratio (Rta) on the speed increasing side before the switching point toroidal ratio (Rtm), the feedback control of the transmission ratio is performed. At the same time, the number of pulses (N) or the sleeve position (S) is fixed to the predetermined number of pulses (Na) and the predetermined sleeve position (Sa) at the time t1, as indicated by the reference sign. The fixing of the number of pulses (N) and the like is continued until a predetermined time Ta elapses from time t1.
[0134]
Here, as shown in FIG. 17, the predetermined toroidal ratio (Rta) is corrected according to the torque (Tr1) passing through the transmission mechanisms 20 and 30 at the time t1, and the torque (Tr1) is zero and the power When the tilting inertia of the rollers 23 and 33 does not occur, the switching point toroidal ratio (Rtm) is set as a theoretical value. However, the torque (Tr1) is increased, and therefore the tilting inertia of the power rollers 23 and 33 is increased. As the value increases, the value is set to a higher speed side. As a result, the power rollers 23 and 33 continue to incline toward the deceleration side due to inertia and then stop. As a result, as shown by the reference symbol, the actual toroidal ratio (Rtr) is obtained at time tx after time t1. The switching point toroidal ratio (Rtm) is reached and stabilized at the ratio (Rtm). The actual toroidal ratio (Rtr) is stably fixed to the switching point toroidal ratio (Rtm) at least until the time point t4 when the normal speed ratio feedback control is resumed after the mode switching is completed. Thus, the pulse control to the forward stepping motor 251 is performed. As a result, a smooth changeover of the clutches 60 and 70 that does not cause significant gear ratio fluctuations and switching shocks is realized.
[0135]
On the other hand, until the actual toroidal ratio (Rtr) reaches the switching point toroidal ratio (Rtm), the number of pulses (N) or the sleeve position (S) continues to be feedback-controlled to the minus side, as indicated by the reference sign, If the control is stopped after the actual toroidal ratio (Rtr) reaches the switching point toroidal ratio (Rtm), the actual toroidal ratio (Rtr) is changed to the switching point toroidal ratio (Rtm) as indicated by the reference symbol “C”. ) Over the deceleration side, and as a result, the gear ratio deviates from the switching point gear ratio, causing an unpleasant shock when the clutches 60 and 70 are switched, or as indicated by the reference symbol The number (Ne) rises unnecessarily.
[0136]
The predetermined number of pulses (Na) and the predetermined sleeve position (Sa) at the time point t1 when the actual toroidal ratio (Rtr) reaches the predetermined toroidal ratio (Rta) are determined based on the transmission torque (Tr1). It can be said that the number of pulses (N) and the sleeve position (S) at which the toroidal ratio (Rta) can be realized. Further, from the target value generated due to the traction force described in FIG. The shift phenomenon is also taken into account.
[0137]
Further, at the time t1, the target toroidal ratio (Rtb) at the time of restarting used in the feedback control of the gear ratio restarted from the time t4 after switching to the high mode HF is calculated. Also in that case, the phenomenon of deviation of the number of pulses from the target value or the sleeve position generated due to the traction force described in FIG. 11 is taken into consideration.
[0138]
On the other hand, at time t1 when the actual toroidal ratio (Rtr) reaches the predetermined toroidal ratio (Rta), the low clutch duty ratio DL is set to 0% (DL2), and the high clutch duty ratio DH is set to the predetermined duty ratio (DH2). Set to As a result, the low clutch 60 starts to be released while the high clutch 70 starts to be engaged. That is, when the actual toroidal ratio (Rtr) reaches the predetermined toroidal ratio (Rta), the changing operation of the clutches 60 and 70 is started.
[0139]
Here, the number of pulses (N) or the sleeve position (S) is fixed to the predetermined number of pulses (Na) and the predetermined sleeve position (Sa) at time t1, and the predetermined time Ta is substantially equal to the clutches 60, 70. With the start of the switching operation, both the low and high clutches 60 and 70 are set to the time required to be in the engaged state as indicated by the reference numeral. When both the low and high clutches 60 and 70 are engaged, the toroidal ratio is maintained at the switching point toroidal ratio (Rtm) regardless of the number of pulses (N) or the sleeve position (S). The Accordingly, as long as the actual toroidal ratio (Rtr) reaches the predetermined toroidal ratio (Rta) at the time t1, the toroidal ratio is stabilized at the switching point toroidal ratio (Rtm) by the start of the clutch 60 and 70 changing operation. Therefore, the feedback control of the number of pulses (N) or the like is no longer performed, so the feedback control is stopped at the time t1, and the number of pulses (N) or the like becomes the predetermined number of pulses (Na). And so on.
[0140]
As shown in FIG. 18, the predetermined time Ta is set to be shorter as the oil temperature is higher. This is because when the oil temperature is high, the viscosity of the hydraulic oil or the lubricating oil is low, the engagement operation of the engagement side friction element (in this case, the high clutch 70) proceeds with good responsiveness, and the actual toroidal ratio (Rtr) is This is because the switching point toroidal ratio (Rtm) is reached quickly. Thereby, the control time required for the mode switching can be shortened.
[0141]
Then, after the time point t2 when the predetermined time Ta elapses, the high clutch duty ratio DH is further increased toward a duty ratio (DH3) of 100%. As a result, the release operation of the low clutch 60 and the engagement operation of the high clutch 70 further proceed from the state where both the low and high clutches 60 and 70 are in the engaged state. That is, the transition from the low mode to the high mode is started.
[0142]
At this time, the feedforward (F / F) control of the number of pulses (N) or the sleeve position (S) is executed while the second predetermined time Tb from the time point t2 to the time point t3 elapses. That is, the torque (Tr1) at time t1 is set as the torque in the low mode LF before mode switching, and the torque (Tr2) at time t2 is set as the torque in the high mode HF after mode switching with reference to FIG. As described above, since the switching point toroidal ratio (Rtm) is maintained and realized in both modes LF and HF, the pulse number (N) for the step motor 251 is corrected by the deviation Δn of the Rtm realization pulse numbers N1 and N2. is there. Here, the characteristic approximation formulas f1 to f4 described above are used to calculate the Rtm realization pulse number. In particular, in this case, the characteristic approximation expression is used to calculate the Rtm realization pulse number N1 in the low mode LF and the positive drive state. The characteristic approximate expression f2 is used for calculating the number of Rtm realization pulses N2 in the high mode HF and the positive drive state.
[0143]
As a result, as indicated by symbol H, the number of pulses (N) is quickly increased by the deviation Δn with good responsiveness by the feedforward control during the second predetermined time Tb. Therefore, the shift to the high mode is started from the state where both the low and high clutches 60 and 70 are in the engaged state. As a result, the power rollers 23 and 33 are accelerated in the speed increasing direction by the traction forces T2 and T2. Even when moving from the offset position to the position offset in the deceleration direction by the traction forces T1 and T1 acting in the opposite direction, as shown by the symbol Z, no significant toroidal ratio or unit ratio fluctuation occurs, The occurrence of an unpleasant shock is suppressed.
[0144]
On the other hand, as indicated by the reference symbol “T”, when the number of pulses (N) is held at the same value as before the start of the shift to the high mode, the difference between the offsets of the power rollers 23 and 33 and the offset direction is caused. As a result, the change of the toroidal ratio in the deceleration direction cannot be avoided, and as a result, the change gear ratio deviates from the change point change gear ratio, or an unpleasant change shock occurs. (Ne) rises unnecessarily here too.
[0145]
In other words, the deviation Δn in the number of pulses is, in other words, feedback control of the switching point toroidal ratio (Rtm) that is the current actual toroidal ratio and the transmission ratio that is restarted from the time t4 after switching to the high mode HF. The difference (Δr) between the target toroidal ratio (Rtb) used is corrected.
[0146]
On the other hand, while the second predetermined time Tb from the time point t2 to the time point t3 has elapsed, the ignition timing (Ig) of the engine 1 is retarded by a predetermined amount ΔIg to reduce the input torque. In this case, as shown in FIG. 19, the retard amount ΔIg is corrected according to the input torque (Tr2) at the time point t2, and as the torque (Tr2) increases, the retard amount ΔIg increases, that is, the torque down amount. Is set to be large. As a result, not only the engagement shock of the high clutch 70 is suppressed, but also the torque fluctuation accompanying the progress of the engagement operation is suppressed, and therefore the amount of offset reversal of the power rollers 23 and 33 is also reduced, and the mode switching shock is further increased. It will be reduced more reliably.
[0147]
Then, after the second predetermined time Tb has elapsed and time t3 has been reached, the feedback control of the number of pulses (N) is still stopped, while the on / off solenoid valve 213 is turned off. As a result, the spool of the fail-safe valve 212 moves to the left, and the low clutch line 274 communicates with the drain port 212b of the valve 212. Therefore, as indicated by the symbol C, the low clutch pressure is quickly discharged from the drain port 212b, and the low clutch 60 is fully opened early and the feedback control of the gear ratio in the high mode HF after switching. Can be executed early.
[0148]
On the other hand, if the on / off solenoid valve 213 remains on, the low clutch pressure is discharged by the duty ratio of the duty solenoid valve 271, and as shown by the symbol H, centrifugal residual pressure and the like remain, It will take time for the high mode HF to be fully realized.
[0149]
Then, when the predetermined value Tc from the time point t2 has elapsed, the high clutch 70 is completely engaged and the low clutch 60 is completely switched to the fully released high mode HF, and the toroidal ratio (Rtb) is targeted. The speed ratio feedback control is resumed.
[0150]
Next, the above control operation will be described with reference to the flowchart of FIG. First, steps S1 to S17 are control operations up to time t1.
[0151]
First, in step S1, various states such as the current vehicle speed, the throttle opening of the engine 1, the selected range, the actual toroidal ratio, the oil temperature, etc. based on the signals from the sensors and switches 301 to 308 shown in FIG. After detecting the amount, the input torque Tr is estimated in step S2. The estimation of the input torque Tr is obtained by a known method from the engine speed, the intake air amount, and the like. Next, at step S3, the engine ignition timing timing Ig is set. This ignition timing timing Ig is also obtained by a known method from the engine speed, the intake air amount, and the like.
[0152]
Next, in step S4, a mode switching determination ratio (predetermined toroidal ratio) Rta is set according to the input torque Tr. In this case, as shown in FIG. 17, the mode switching determination ratio Rta is obtained as a smaller value (a value on the acceleration side) as the input torque Tr is larger.
[0153]
Next, at step S5, it is determined whether or not the actual toroidal ratio Rtr is larger than the mode switching determination ratio Rta. As a result, when it is not large, that is, when the actual toroidal ratio Rtr is higher than the mode switching determination ratio Rta, the routine proceeds to step S6, where the value of the input torque Tr is set to the first torque value Tr1.
[0154]
Next, in step S7, the target value Rto of the toroidal ratio is set to the mode switching theoretical value (switching point toroidal ratio) Rtm, and in steps S8 and S9, the high clutch duty ratio DH is set to the first duty ratio DH1 (0%), and the low clutch duty value is set. DL is set to the first duty ratio DL1 (100%), and the instruction value for the on / off solenoid valve 213 is turned on in step S10.
[0155]
Next, in step S11, a target unit ratio Ruo or a target toroidal ratio Rto is first set based on the shift map, and then a deviation ΔRt of the actual toroidal ratio Rtr with respect to the target toroidal ratio Rto is calculated.
[0156]
In step S12, a proportional term gain (feedback gain in gear ratio control) G in the index Q of PID control is set according to the input torque, gear ratio, travel mode, and range. The proportional term gain G is set to a smaller value as the torque is larger and the toroidal ratio is closer to the GN ratio (Rtn). Further, when the driving mode is the low mode LF or LR, it is set to a smaller value than when the driving mode is the high mode HF, and when the range is the reverse range Lr, the value is smaller than when the driving range is the forward range. Is set.
[0157]
In step S13, the PID control index Q is first calculated by substituting the proportional term gain G into the following equation.
[0158]
[Equation 5]

Here, H is an integral term gain, and I is a differential term gain.
[0159]
Next, based on a map set in advance as shown in FIG. 22, the pulse number deviation ΔN of the control signal output to the step motors 251 and 252 corresponding to the calculated index Q is obtained.
[0160]
Here, in the above map, the pulse number deviation ΔN is set to increase as the absolute value of the index Q increases. When the index Q is positive (when the actual toroidal ratio is larger than the target toroidal ratio), the pulse number deviation ΔN is set to be positive, and the pulse number N is increased, thereby reducing the actual toroidal ratio. When the index Q is negative (when the actual toroidal ratio is smaller than the target toroidal ratio), the pulse number deviation ΔN is set to be negative and the pulse number N is decreased, thereby increasing the actual toroidal ratio. (See FIG. 8).
[0161]
In steps S14 to S17, the control operation up to the time point t1 as described above is obtained by controlling the ignition timing of the step motor 251, the duty solenoid valves 271, 272, the onoff solenoid valve 213, and the engine 1, respectively. .
[0162]
On the other hand, when the actual toroidal ratio Rtr reaches the mode switching determination ratio Rta in step S5, the process proceeds from step S5 to step S18. In this case, the operations from steps S1 to S5, S18 to S24, and S14 to S17 are control operations from time t1 to time t2.
[0163]
The first timer tim1 is set only immediately after the actual toroidal ratio Rtr reaches the mode switching determination ratio Rta in step S18. In step S19, it is determined whether or not the timer tim1 has passed the first predetermined period Ta. As a result, when the first predetermined period Ta has not yet passed, that is, until the time point t2, the value of the input torque Tr is set to the second torque value Tr2 in step S20.
[0164]
In step S21, 0 is substituted as the pulse number deviation ΔN for the step motor 251, and then in steps S22 and S23, the high clutch duty ratio DH is set to the second duty ratio DH2, which is a predetermined shelf pressure, and the low clutch duty is set. The rate DL is set to a second duty rate DL2 that is 0%. In step S24, the instruction value for the on / off solenoid valve 213 is turned on.
[0165]
In steps S14 to S17, the predetermined time Ta elapses from the time point t1 as described above by controlling the ignition timing of the step motor 251, the duty solenoid valves 271, 272, the onoff solenoid valve 213, and the engine 1, respectively. A control operation up to time t2 is obtained.
[0166]
On the other hand, when the first timer tim1 has passed the first predetermined period Ta in step S19, the process proceeds from step S19 to S25. In this case, the operations from steps S1 to S5 and S18, S19 and S25 to S32, and S14 to S17 are control operations from time t2 to time T3.
[0167]
In step S25, only after the first predetermined time Ta has elapsed, the first and second torque values Tr1 and Tr2 are substituted into the characteristic approximation equations f1 and f2, respectively, and the pulse numbers N1 and N2 and their deviations are substituted. Δn is obtained. Next, in step S26, the second timer tim2 is set only immediately after the first predetermined time Ta has elapsed.
[0168]
In step S27, it is determined whether or not the second timer tim2 has passed the second predetermined time Tb. If it has not passed yet, that is, from time t2 to time t3, the process proceeds to step S28. The pulse number deviation ΔN of the output signal to the step motor 251 is set so that the sleeve 222 of the three-layer valve 220 moves by the pulse number deviation Δn during the second predetermined time Tb.
[0169]
Next, in step S29, the instruction value for the on / off solenoid valve 213 is turned on, and in step S30, the ignition timing Ig is retarded by a predetermined value ΔIg to reduce the torque.
[0170]
Next, in step S31, the low clutch duty ratio DL is changed to the second duty ratio DL2, and in step S32, the high clutch duty ratio DH is set between the second predetermined time Tb and the next third predetermined time. The ratio DH is set to a value that changes from the second duty ratio DH2 to the third duty ratio DH3. Here, the value of the third duty factor DH3 is 100%.
[0171]
In steps S14 to S17, the predetermined time Tb elapses from the time point t2 as described above by controlling the ignition timing of the step motor 251, the duty solenoid valves 271, 272, the onoff solenoid valve 213, and the engine 1, respectively. A control operation up to time t3 is obtained.
[0172]
On the other hand, when it is determined in step S27 that the second predetermined time Tb has elapsed, the process proceeds from step S27 to S33, where the second timer tim2 is set to the second predetermined time Tb and the third predetermined time Tc. It is determined whether or not the sum is greater than or equal to. If NO, that is, between time t3 and time T4, the process proceeds to step S34, where the pulse number deviation ΔN is set to 0 again, and OFF is set as an instruction signal for the on / off solenoid valve in step S35. The
[0173]
By the above control, the control operation from the time t3 as described above to the time t4 when the predetermined time Tc elapses is obtained.
[0174]
Then, after the third predetermined time Tc has elapsed and time t4 has been reached, NO is determined in step S33, and the process proceeds from step S33 to S36. In step S36, the third duty ratio DH3 having a high clutch duty ratio DH of 100% is substituted, and in step S37, the low clutch duty ratio DL is maintained at a second duty ratio DL2 of 0%.
[0175]
As a result, the control operation after the time t4 when the high clutch 70 is completely engaged and the mode is completely switched to the high mode is obtained, and the normal feedback control of the gear ratio is resumed, and the gear ratio corresponding to the running state is resumed. Control will be executed.
[0176]
In the above description, the first predetermined time Ta is corrected by the oil temperature, and this is measured by the first timer tim1 in step S19. Instead, the low mode clutch 60 and the high mode clutch 70 are used. In order to determine whether or not both are in the engaged state, the time point when the hydraulic pressure of the engagement side friction element becomes higher than a predetermined value may be set as t2.
[0177]
Furthermore, when both the clutches 60 and 70 are engaged and the rotational slip amount between the output side rotating members 60a and 60b; 70a and 70b (see FIG. 1) is less than a predetermined value, the sleeve 222 of the three-layer valve 220 is provided. May be moved by the pulse number deviation Δn.
[0178]
In that case, as shown in FIG. 1 and FIG. 7, an output shaft rotational speed sensor 310 that detects the rotational speed of the secondary shaft 13 is provided, and this sensor 310 detects the rotational speed of the secondary shaft 13 and The rotation speed of the output disks 21 and 22 is detected by the input rotation speed sensors 307 and 308.
[0179]
For example, the slip amount SLP (60) of the low clutch 60 is expressed by the following equation from the rotational speed W (60a) of the input side rotating member 60a and the rotational speed W (60b) of the output side rotating member 60b. .
[0180]
[Formula 6]

Here, the rotational speed W (60a) of the input side rotating member 60a of the low clutch 60 is the rotational speed W (21) of the input disk 21 and the number of teeth Z of the first gear 81 and the second gear 82 of the low mode gear train 80. (81) and Z (82) can be calculated according to the following equation.
[0181]
[Expression 7]

Further, the rotational speed W (60b) of the output side rotating member 60b of the low clutch 60 includes the rotational speed W (22) of the output disk 22 and the number of teeth Z (52 of the sun gear 52 and the internal gear 53 of the planetary gear mechanism 50). ), Z (53), the number of teeth Z (91), Z (92) of the first gear 91 and the second gear 92 of the high mode gear train 90, and the rotational speed W (13) of the secondary shaft 13 It can be calculated according to the formula.
[0182]
[Equation 8]

Since the number of teeth Z is known in advance from the specifications, the slip amount SLP (60) of the low clutch 60 is eventually determined by the rotational speed W (13) of the secondary shaft 13 and the rotational speeds of the input and output disks 21 and 22. It can be calculated from W (21) and W (22).
[0183]
On the other hand, similarly, the slip amount SLP (70) of the high clutch 70 is calculated from the rotational speed W (70a) of the input side rotating member 70a and the rotational speed W (70b) of the output side rotating member 70b as shown in the following equation. expressed.
[0184]
[Equation 9]

Here, the rotation speed W (70a) of the input side rotation member 70a of the high clutch 70 is the rotation speed W (22) of the output disk 22, and the teeth of the first gear 91 and the second gear 92 of the high mode gear train 90. It can be calculated according to the following equation from the numbers Z (91) and Z (92).
[0185]
[Expression 10]

The rotational speed W (70b) of the output side rotating member 70b of the high clutch 70 is the rotational speed W (13) of the secondary shaft 13, and the number of teeth Z is known in advance from the specifications. The slip amount SLP (70) of the clutch 70 can be calculated from the rotational speed W (13) of the secondary shaft 13 and the rotational speed W (22) of the output disk 22.
[0186]
【The invention's effect】
As described above, according to the present invention, in a power train provided with a continuously variable transmission mechanism, when the power transmission path is switched, fluctuations in the gear ratio accompanying the switching are suppressed, so that the power train is stable before and after switching. Thus, the gear ratio can be maintained at the target gear ratio, and the occurrence of unpleasant shock when switching the travel mode is effectively eliminated. The present invention can be preferably applied to, for example, a power train including a toroidal-type continuously variable transmission mechanism, and can be widely used in the vehicle industry in general.
[0187]
[Prior art]
[Brief description of the drawings]
FIG. 1 is a skeleton diagram showing a mechanical configuration of a toroidal continuously variable transmission according to an embodiment of the present invention.
FIG. 2 is a plan view showing a specific structure of a main part of the transmission in an unfolded state.
FIG. 3 is a cross-sectional view taken along line AA in FIG.
FIG. 4 is a circuit diagram of hydraulic control of the transmission.
5 is a partial cross-sectional view of the periphery of a shift control three-layer valve as viewed from the direction B in FIG. 3;
6 is a partial cross-sectional view of the periphery of the speed change control mechanism as viewed from the direction C in FIG.
FIG. 7 is a block diagram showing a control system for the entire power train.
FIG. 8 is a characteristic diagram showing a relationship between the number of pulses of a step motor and a toroidal ratio.
FIG. 9 is a characteristic diagram showing the relationship between the number of pulses of the step motor and the unit ratio.
FIG. 10 is an explanatory diagram of circulating torque.
FIG. 11 is a characteristic diagram showing the relationship between the number of pulses of a step motor and torque.
FIG. 12 is a characteristic diagram showing a relationship between a unit ratio and a toroidal ratio.
FIG. 13 is a shift diagram used in the forward low mode.
FIG. 14 is a shift diagram used in the forward high mode.
FIG. 15 is a shift diagram used in a reverse low mode.
FIG. 16 is a time chart of the switching control operation from the low mode to the high mode in the positive drive state.
FIG. 17 is a map showing a relationship between a torque used in the control and a toroidal ratio for determining mode switching start.
FIG. 18 is a map showing a relationship between an oil temperature used in the control and a first predetermined time.
FIG. 19 is a map showing the relationship between torque and retard amount used in the control.
FIG. 20 is a flowchart showing a part of the control operation.
FIG. 21 is a flowchart showing a part of the same.
FIG. 22 is a map showing a relationship between an index used in the control operation and the number of pulses of a control signal output to the step motor.
FIG. 23 is a flowchart showing a part of the control operation.
FIG. 24 is a flowchart showing a part of the same.
FIG. 25 is a flowchart showing a part of the same.
[Explanation of symbols]
1 engine
10 Powertrain
20, 30 continuously variable transmission mechanism
50 Planetary gear mechanism (gear mechanism)
60 Low clutch (friction element)
70 High clutch (friction element)
200 Hydraulic control circuit
251,252 Step motor (speed ratio control means)
271 Low clutch duty solenoid valve (path switching means)
272 Duty solenoid valve for high clutch (path switching means)
300 Control unit (switching control means)

Claims (7)

A first path passing through the continuously variable transmission mechanism and the gear mechanism and a second path passing only through the continuously variable transmission mechanism are provided, and path switching means for selectively switching these paths is provided. In addition, the continuously variable transmission mechanism includes an input disk and an output disk, a power roller interposed between the two disks, and a trunnion that supports the power roller. Because
Gear ratio control means for controlling the gear ratio of the continuously variable transmission mechanism by controlling the tilt angle of the power roller with respect to the input disk and the output disk by controlling the position of the trunnion with respect to the input disk and the output disk ;
Torque detecting means for detecting torque;
Switching control means for controlling a position control amount of the trunnion by the speed ratio control means so as to suppress a change in speed ratio due to the switching when the path is switched by the path switching means ;
The power control apparatus, wherein the switching control means increases the correction amount of the trunnion position control amount toward the deceleration side or the acceleration side as the torque detected by the torque detection means increases . The switching control means reverses the direction of correction of the trunnion position control amount at the time of switching from the first path to the second path and at the time of switching from the second path to the first path. The powertrain control device according to claim 1. The switching control means determines the direction of correction of the trunnion position control amount when the vehicle is in a normal drive state driven by an engine and when the engine is in a reverse drive state driven by the running inertia of the vehicle. powertrain control device according to claim 1 or 2, characterized in that reversing. Provided friction element to achieve the first path by being fastened respectively and the friction element to achieve the second path, said path switching means, either through both a state of being entered into both of these frictional elements or with one of the friction elements are configured to perform the switching of paths by selectively fastening, the switching control means, after a predetermined time has elapsed since the start of a change over by that path to the path switching means powertrain control system according to any one of claims 1 to 3, characterized by controlling the amount of position control of the trunnion by the speed ratio control means. Said switching control means, when the fastening pressure of the engagement side frictional element becomes a predetermined value or more, according to claim 4, characterized in that for controlling the position control amount of the trunnion by the gear ratio control means Powertrain control device. The switching control means is configured to change the transmission ratio control means when the engagement-side friction element is inserted, the differential rotation of the output rotation member, and the release-side friction element is input, and the differential rotation of the output rotation member is less than a predetermined value. The power train control device according to claim 4 , wherein the position control amount of the trunnion is controlled by . Provided with a torque decrease means for decreasing the torque, the switching control means, the power of claim 4 where the torque at the time of engagement of the engagement side frictional element and controls the torque reduction means so as to reduce train Control device.

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