JP3892510B2 - Vehicle left / right driving force distribution device - Google Patents

Description

同様にして、
Ｆｓ2 ＝Ｐ2 ・ｔａｎα2 ／ｃｏｓβ2
Ｆｔ2 ＝Ｐ2 ・ｔａｎβ2
が成立して、プラネタリンピン４３側に作用する合成力Ｎｐ2 は以下のようになる。

また、第１，第２のピニオン４１，４２内に生じる残留スラスト力ΔＦｔは以下のようになる。

従って摩擦トルクＴｆは、２つの合成力Ｎｐ1 ，Ｎｐ2 による摩擦力、残留スラスト力ΔＦｔによる摩擦力との和で、以下のようになる。

次いで、第１，第２のピニオン４１，４２でのトルクのバランス式は、以下のようになる。
Ｔｆ＋Ｐ・ｒｐ1 ＝Ｐ2 ・ｒｐ2 …（１５）
また、上記（１０）式に摩擦トルクＴｆ分を加えると、以下のようになる。

ここで、前述のように、上記各かみ合いピッチ円半径ｒｓ1 ，ｒｓ2 ，ｒｐ1 ，ｒｐ2 を各歯車の歯数Ｚｓ1 ，Ｚｓ2 ，Ｚｐ1 ，Ｚｐ2 で置き換え、これら各歯数に前記各歯数を代入する（Ｚｐ1 ＝２４，Ｚｐ2 ＝２４，Ｚｓ1 ＝３０，Ｚｓ2 ＝１５）と、上記（１６）式は、
ＴＢ＝０．５Ｔｉ＋０．６２５Ｔｆ …（１７）
となる。
【００６４】
また、Ｔｉ＝ＴＦ＋ＴＢであり、これに上記（１６）式を代入して整理すると、以下のようになる。

さらに、各歯数Ｚｓ1 ，Ｚｓ2 ，Ｚｐ1 ，Ｚｐ2 で置き換え、これら各歯数に前記各歯数を代入すると、上記（１８）式は、
ＴＦ＝０．５Ｔｉ−０．６２５Ｔｆ …（１９）
となる。
【００６５】
ここで、μ1 ＝０，μ2 ＝０なら、Ｔｆ＝０であり、前後輪側トルクＴＦ，ＴＢの値は、上述の等トルク配分機能の場合の式と同一の基準トルク配分を示す。
【００６６】
こうして、かかる条件では、摩擦トルクＴｆに応じた差動制限トルクＴｆ・ｒｓ2 ／ｒｐ2 が発生することがわかる。そして前後輪側トルクＴＦ，ＴＢの配分が、差動制限トルクの分だけ、後輪側が大きく、前輪側が小さくなるように変化する。また、摩擦トルクＴｆが生じる合成力Ｎｐ1 ，Ｎｐ2 ，残留スラスト力ΔＦｔは入力トルクに比例するため、入力トルク比例式差動制限機能を有する。
【００６７】
一方、第１，第２のピニオン４１，４２のねじれ角β1 とβ2 との差により残留スラスト力ΔＦｔが変えられ、また、上記プラネタリピン４３の接触摩擦部分にニードルベアリングやブッシュ等を用いることにより、摩擦係数μ1 を変えることができる。このように、摩擦トルクＴｆとともに差動制限トルクの値を様々な値に定めることが可能になっている。
【００６８】
続いて、ＮＢ＞ＮＦの場合について説明する。この条件では、図９のようになり、第１，第２のピニオン４１，４２が第１のサンギヤ３９と反対の時計方向に自転しながら公転して、摩擦トルクＴｆは反時計方向に作用する。このため、第１，第２のピニオン４１，４２内のトルクのバランス式は以下のようになる。
Ｔｆ＋Ｐ2 ・ｒｓ2 ＝Ｐ・ｒｐ1 …（２０）
そして上述と同様に計算すると、前後輪側トルクＴＦ，ＴＢは以下のようになる。

従ってこの条件でも同一の差動制限トルク、Ｔｆ・ｒｓ2 ／ｒｐ2 が発生する。一方、この場合は上述と逆に差動制限トルク分だけ後輪側が小さく、前輪側が大きくなるようにトルク配分されことになる。
【００６９】
次に、前記後輪左右駆動力配分部７について、図２を基に詳しく説明する。
上記後輪左右駆動力配分部７は、大きく分けて差動制限機構部４４と、歯車機構部４５と、クラッチ機構部４６とから主に構成されており、この後輪左右駆動力配分部７に駆動力を伝達する前記ドライブピニオン６と上記差動制限機構部４４は、ディファレンシャルキャリア４７内に収容され、上記クラッチ機構部４６は上記歯車機構部４５を介して上記ディファレンシャルキャリア４７側面に配設され、このディファレンシャルキャリア４７の後端部はカバー４８で覆われ、上記歯車機構部４５と上記クラッチ機構部４６はカバー４９で覆われている。
【００７０】
上記ドライブピニオン６は、前記プロペラシャフト５に接続する軸部６ａが、上記ディファレンシャルキャリア４７内に軸受で回転自在に支持されている。
【００７１】
まず、上記差動制限機構部４４について説明する。
前記左ドライブ軸１４は上記ディファレンシャルキャリヤ４７に取り付けられた左側サイドリテーナ５０を貫通して回転自在に設けられ、この左ドライブ軸１４と同軸上に、前記右ドライブ軸１７が上記ディファレンシャルキャリヤ４７の右側を貫通して回転自在に設けられている。
【００７２】
上記左ドライブ軸１４の外周には、左側ディファレンシャルケース５１Ｌが回転自在に嵌合され、上記左ドライブ軸１４と上記左側ディファレンシャルケース５１Ｌが、上記左側サイドリテーナ５０に軸受を介して回転自在に支持されている。
【００７３】
上記左側ディファレンシャルケース５１Ｌには、右側ディファレンシャルケース５１Ｒの一端部（左側端部）と、上記ドライブピニオン６と噛合されるクラウンギヤ５２が共に回転中心の芯合わせがなされ固定されており、上記右側ディファレンシャルケース５１Ｒの他端部５１Ｒａは円筒状に形成されて、キャリヤ５３の右側面に形成した筒部５３ａの外周に回転自在に嵌合されるとともに、上記歯車機構部４５における第１の歯車軸５４の外周にスプライン嵌合されて軸受を介して上記ディファレンシャルキャリア４７に支持されている。
【００７４】
また、上記キャリヤ５３の筒部５３ａの内側には、上記歯車機構部４５における第２の歯車軸５５がスプライン嵌合され、さらに、この第２の歯車軸５５の内側に上記右ドライブ軸１７が回転自在に嵌合されている。
【００７５】
すなわち、上記右側ディファレンシャルケース５１Ｒと上記第１の歯車軸５４の接続部と、上記キャリヤ５３と上記第２の歯車軸５５の接続部と、上記右ドライブ軸１７とは、それぞれが回転自在に上記ディファレンシャルキャリア４７に支持され、また、上記左側ディファレンシャルケース５１Ｌと上記右側ディファレンシャルケース５１Ｒとで構成され、上記クラウンギヤ５２が取り付けられたディファレンシャルケース５１が、上記ディファレンシャルキャリア４７内で回転自在に保持されている。
【００７６】
上記ディファレンシャルケース５１内には、上記キャリヤ５３が回転自在に配設されて、上記キャリヤ５３内には上記左ドライブ軸１４と上記右ドライブ軸１７が挿入されて、上記キャリヤ５３が上記左ドライブ軸１４の先端部にスプライン結合されている。
【００７７】
上記ディファレンシャルケース５１内で、上記左側ディファレンシャルケース５１Ｌの上記左ドライブ軸１４が挿通される部分には、大径の第１のサンギヤ５６がスプライン結合され、上記右ドライブ軸１７の先端部には小径の第２のサンギヤ５７がスプライン結合され、上記第１のサンギヤ５６が小径の第１のピニオン５８と噛合して第１の歯車列が形成され、上記第２のサンギヤ５７が大径の第２のピニオン５９と噛合して第２の歯車列が形成されている。
【００７８】
上記第１のピニオン５８と上記第２のピニオン５９はピニオン部材６０に一体に形成されており、複数（例えば３個）の上記ピニオン部材６０が、キャリヤ５３に固定したそれぞれのプラネタリピン６１に軸受を介して回転自在に軸支されている。上記ピニオン部材６０の両端には上記キャリア５３との間にスラスト荷重受け用のワッシャが介装されている。
【００７９】
すなわち、上記差動制限機構部４４は、上記ドライブピニオン６からの駆動力を、クラウンギヤ５２，ディファレンシャルケース５１を介して第１のサンギヤ５６に伝達し、上記第２のサンギヤ５７から上記右ドライブ軸１７へ出力する一方、上記キャリヤ５３から上記左ドライブ軸１４へ出力する複合プラネタリギヤ式の差動制限装置で構成されている。
【００８０】
上記差動制限機構部４４が形成する上記複合プラネタリギヤ式の差動制限装置では、上記第１，第２のサンギヤ５６，５７およびこれらサンギヤ５６，５７の周囲に複数個配置される上記第１，第２のピニオン５８，５９の歯数を適切に設定することで差動機能を有する。
【００８１】
また、上記第１，第２のサンギヤ５６，５７と上記第１，第２のピニオン５８，５９とのかみ合いピッチ円半径を適切に設定することで、基準トルク配分が左右５０：５０の等トルク配分の機能を有する。
【００８２】
更に、上記第１，第２のサンギヤ５６，５７と上記第１，第２のピニオン５８，５９とを例えばはすば歯車にし、上記第１の歯車列と上記第２の歯車列のねじれ角を異にしてスラスト荷重を相殺させることなくスラスト荷重を残留させピニオン端面間に摩擦トルクを、上記第１，第２のピニオン５８，５９と上記プラネタリピン６１の表面に噛合いによる分離，接線荷重の合成力が作用し、摩擦トルクが生じるように設定して、入力トルクに比例した差動制限トルクを得られるようにすることで差動制限機能を有する。
【００８３】
尚、上記差動制限機構部４４が形成する上記複合プラネタリギヤ式の差動制限装置のさらに詳しい説明は、前記センターディファレンシャル装置３での説明と略同様であるので省略する。
【００８４】
次に、上記歯車機構部４５について説明する。
上記ディファレンシャルケース５３と連結された上記第１の歯車軸５４は、上記ディファレンシャルキャリア４７の外側に延出されて軸受を介して上記ディファレンシャルキャリア４７に回転自在に支持され、上記キャリヤ５３との接続部とは逆の端部に、第１の歯車６２が形成されている。
【００８５】
また、上記キャリヤ５３と連結された上記第２の歯車軸５５は、上記第１の歯車軸５４内を回転自在に上記ディファレンシャルキャリア４７の外側に延出され、端部に上記第１の歯車６２の外側（右側）に隣接して配設された第２の歯車６３がスプライン結合されている。
【００８６】
さらに、上記第２の歯車６３の外側には、上記右ドライブ軸１７にスプライン結合された第３の歯車６４が設けられている。
【００８７】
そして、上記各第１，２，３の歯車６２，６３，６４は、それぞれ、上記第１，２，３の歯車６２，６３，６４の回転軸芯と平行な同一回転軸芯上に並設された第４，５，６の歯車６５，６６，６７と噛合されている。
【００８８】
すなわち、上記歯車機構部４５は、上記第１の歯車６２と上記第４の歯車６５による第１の歯車列と、上記第２の歯車６３と上記第５の歯車６６による第２の歯車列と、第３の歯車６４と第６の歯車６７による第３の歯車列の３つの歯車列から構成されている。
【００８９】
そして、上記各歯車列のそれぞれのギヤ比は、上記第１，２，３，４，５，６の歯車６２，６３，６４，６５，６６，６７の歯数をそれぞれｚ1 ，ｚ2 ，ｚ3 ，ｚ4 ，ｚ5 ，ｚ6 として、第１の歯車列は、ｚ4 ／ｚ1 ＝０．９、第２の歯車列は、ｚ5 ／ｚ2 ＝１、第３の歯車列は、ｚ6 ／ｚ3 ＝１に設定され、第１の歯車列のギヤ比だけが増速比に設定されている。
【００９０】
上記第４の歯車６５は、第４の歯車軸６８の一端側に一体に形成され、上記第４の歯車軸６８の一端部が軸受を介して上記ディファレンシャルキャリア４７の外側に回転自在に支持されて第１の回転部材となっており、他端部が上記カバー４９に回転自在に支持されている。また、上記第４の歯車軸６８の他端側には、上記第４の歯車６５側が開口した円筒状のクラッチドラム６９が固定されている。
【００９１】
また、上記第５の歯車６６は、第５の歯車軸７０の一端側に一体に形成されて第２の回転部材となっており、この第５の歯車軸７０は上記第４の歯車軸６８の外周に沿って上記クラッチドラム６９の底部６９ａまで延出して設けられている。上記第５の歯車軸７０の他端部には所定長さのクラッチハブ７１が設けられ、このクラッチハブ７１と上記クラッチドラム６９の間に複数プレート７２を交互に重ねて設け、第１の油圧多板クラッチ７３が形成されている。
【００９２】
さらに、上記第６の歯車６７は、第６の歯車軸７４の一端側に一体に形成されて第３の回転部材となっており、この第６の歯車軸７４は上記第５の歯車軸７０の外周に沿って上記第１の油圧多板クラッチ７３のプレート端面まで延出して設けられている。上記第６の歯車軸７４の他端部には所定長さのクラッチハブ７５が形成され、このクラッチハブ７５と上記クラッチドラム６９の間に複数プレート７６を交互に重ねて設けて第２の油圧多板クラッチ７７が形成されている。
【００９３】
上記第１の油圧多板クラッチ７３は、上記クラッチドラム底部６９ａから挿通された第１のピストン７８により押圧自在になっており、この第１のピストン７８を動作させる第１の油圧室７９は前記油圧管路３６の第１の油圧管路３６ａと連通されている。そして、上記第１のピストン７８を動作させるために、上記第１の油圧室７９に加えられる設定油圧は、前記左右駆動力配分制御部３２により制御された値で可変になっている。
【００９４】
同様に、上記第２の油圧多板クラッチ７７は、上記クラッチドラム底部６９ａから挿通された第２のピストン８０により押圧自在になっており、この第２のピストン８０を動作させる第２の油圧室８１は前記油圧管路３６の第２の油圧管路３６ｂと連通されている。そして、上記第２のピストン８０を動作させるために、上記第２の油圧室８１に加えられる設定油圧は、前記左右駆動力配分制御部３２により制御された値で可変になっている。
【００９５】
すなわち、上記クラッチ機構部４６は、上記２つのクラッチ７３，７７を一体に形成したものであり、前述したように上記歯車機構部４５のギヤ比等が設定されていることから、上記第１の油圧多板クラッチ７３を連結させると上記キャリヤ５３側の左ドライブ軸１４に駆動力が多く配分され、上記第２の油圧多板クラッチ７７を連結させると上記右ドライブ軸１７に駆動力が多く配分されるようになっている。ここで、上記各油圧多板クラッチ７３，７７を連結させる油圧値は上記左右駆動力配分制御部３２によって演算された値であり、この油圧値の大小によってトルク配分量が変化されるのである。
【００９６】
さらに、上記歯車機構部４５と上記クラッチ機構部４６で行われる駆動力配分について説明する。まず、上記右ドライブ軸１７に駆動力が多く配分されるようにして左旋回性能を向上させるには、上記右ドライブ軸１７側の第２の油圧多板クラッチ７７を設定油圧で連結する。上記キャリヤ５３側（左ドライブ軸１４側）に配分される駆動力と右ドライブ軸１７側に配分される駆動力は、上記差動制限機構部４４が動作しなければ、５０対５０の関係にあるので、後輪左右駆動力配分部７に入力される駆動力をＴＢとすると、それぞれＴＢ／２となる。また、上記第２の油圧多板クラッチ７７から上記右ドライブ軸１７に多く配分される駆動力は、第３の歯車列を経て配分されることになる。このため、上記第２の油圧多板クラッチ７７に関し、上記第２の油圧室８１に作動する油圧、摩擦面の動摩擦係数（摩擦面の相対回転速度で決まる動摩擦係数）、摩擦面の枚数（多板クラッチの枚数×２）、有効半径等により決定されるクラッチのスリップトルクをＴｋ２とすると、上記右ドライブ軸１７に配分される駆動力（右輪側駆動力）と上記キャリヤ５３側に配分される駆動力（左輪側駆動力）は、次式のようになる。
右輪側駆動力＝ＴＢ／２＋（Ｔｋ２／２×（ｚ3 ／ｚ6 ））×ＴＢ…（２５）
左輪側駆動力＝ＴＢ／２−（Ｔｋ２／２×（ｚ3 ／ｚ6 ））×ＴＢ…（２６）
そして、上記（２５），（２６）式に前記ギヤ比ｚ6 ／ｚ3 ＝１を代入して、
右輪側駆動力＝ＴＢ／２＋Ｔｋ２×（ＴＢ／２） …（２５）'
左輪側駆動力＝ＴＢ／２−Ｔｋ２×（ＴＢ／２） …（２６）'
となる。
【００９７】
また、上記キャリヤ５３側に駆動力が多く配分されるようにして右旋回性能を向上させるには、上記キャリヤ５３側の上記第１の油圧多板クラッチ７３を設定油圧で連結する。上記第１の油圧多板クラッチ７３から上記キャリヤ５３に多く配分される駆動力は、第２の歯車列を経て配分されることになる。このため、上記第１の油圧多板クラッチ７３に関し、上記第１の油圧室７９に作動する油圧、摩擦面の動摩擦係数（摩擦面の相対回転速度で決まる動摩擦係数）、摩擦面の枚数（多板クラッチの枚数×２）、有効半径等により決定されるクラッチのスリップトルクをＴｋ１とすると、上記右ドライブ軸１７に配分される駆動力（右輪側駆動力）と上記キャリヤ５３側に配分される駆動力（左輪側駆動力）は、次式のようになる。
右輪側駆動力＝ＴＢ／２−（Ｔｋ１／２×（ｚ2 ／ｚ5 ））×ＴＢ…（２７）
左輪側駆動力＝ＴＢ／２＋（Ｔｋ１／２×（ｚ2 ／ｚ5 ））×ＴＢ…（２８）
そして、上記（２７），（２８）式に前記ギヤ比ｚ5 ／ｚ2 ＝１を代入して、
右輪側駆動力＝ＴＢ／２＋Ｔｋ１×（ＴＢ／２） …（２７）'
左輪側駆動力＝ＴＢ／２−Ｔｋ１×（ＴＢ／２） …（２８）'
となる。
【００９８】
上記クラッチ機構部４６の２つのクラッチは、上述の油圧多板クラッチ以外に電磁クラッチや伝達容量可変なカップリングを用いても良い。
【００９９】
一方、前記前輪左右駆動力配分部１２は、駆動力が前記フロントドライブ軸１０，ドライブピニオン１１からクラウンギヤ５２に入力されるようになっており、その構造は上記後輪左右駆動力配分部７と略同様であるので説明は省略する。
【０１００】
次いで、前記後輪側油圧制御装置３４および前記前輪側油圧制御装置３５について、図１０を基に説明する。上記後輪側油圧制御装置３４は、制御油圧を油圧管路３６ａを通じて第１の油圧室７９に加える油圧経路と、油圧管路３６ｂを通じて第２の油圧室８１に加える油圧経路の一対の油圧経路を備えて構成され、上記前輪側油圧制御装置３５も略同様に構成されている。このため、制御油圧を油圧管路３６ａを通じて第１の油圧室７９に加える油圧経路についてのみ以下説明する。
【０１０１】
モータ８２により駆動されるオイルポンプ８３の吐出圧がレギュレータ弁８４で調圧され、所定の作動油圧と潤滑油圧を生じるようになっており、作動油圧の油路８５は、クラッチ制御弁８６，油圧管路３６ａを介して第１の油圧多板クラッチ７３の前記第１の油圧室７９側に連通されている。
【０１０２】
また、上記油路８５は、パイロット弁８７，油路８８によりデューティソレノイド弁８９，上記クラッチ制御弁８６の制御側に連通されている。
【０１０３】
そして、前記左右駆動力配分制御部３２からのデューティ信号は、上記デューティソレノイド弁８９に出力されてデューティ圧が生じ、このデューティ圧で上記クラッチ制御弁８６を動作することで、上記第１の油圧多板クラッチ７３のクラッチ油圧を制御するようになっている。
【０１０４】
また、上記左右駆動力配分制御部３２は、図１１に示すように、路面・走行状態判断部９０、油圧演算部９１、油圧設定部９２から主に構成されており、路面・走行状態に応じて、前後の左右輪間の最適な駆動力配分量を演算し、前後輪側の上記各油圧制御装置３５，３４に信号出力するようになっている。
【０１０５】
上記路面・走行状態判断部９０は、前記スロットル開度センサ２６，車速センサ２７，舵角センサ２９，前後加速度センサ３０，横加速度センサ３１、およびギヤ位置信号が入力され、これらの信号に基づき路面状況（低μ路走行状態か否か等）と走行状態（高速か低速か・急旋回か否か・高負荷か低負荷か（加速状態か）・スリップ状態の有無等）を、予めメモリしておいたマップ、計算式等により求め上記油圧演算部９１に出力するように形成されている。
【０１０６】
また、上記油圧演算部９１では、上記路面・走行状態判断部９０からの信号を基に、予めメモリしておいたマップ、計算式等により、動作させる油圧多板クラッチの選択と、それに付加する油圧値とを演算して、この選択・演算の結果を上記油圧設定部９２に出力するようになっている。
【０１０７】
そして、上記油圧設定部９２は、上記油圧演算部９１からの信号を、それぞれ該当する油圧制御装置に対して信号出力するように形成されている。
【０１０８】
次いで、上記構成の作用を説明する。
先ず、エンジン１による駆動力は、自動変速装置２からトランスミッション出力軸２ａを経てセンターディファレンシャル装置３の第１のサンギヤ３９に入力される。
【０１０９】
そして第１，第２のピニオン４１，４２から第２のサンギヤ４０と、第１，第２のピニオン４１，４２を支持するキャリヤ３８とに分配されて伝達し、上記第２のサンギヤ４０の動力は、リヤドライブ軸４を介して後輪側に伝達される。また、上記キャリヤ３８の動力は、トランスファドライブギヤ８，トランスファドリブンギヤ９，フロントドライブ軸１０を介して前輪側に伝達され４輪駆動で走行する。
【０１１０】
そこで、例えば前輪側回転数と後輪側回転数が等しいＮＦ＝ＮＢの直進走行では、センターディファレンシャル装置３において上記第２のサンギヤ４０と上記キャリヤ３８とが同一方向に等速回転することで、上記第１，第２のピニオン４１，４２は遊星回転しなくなり一体化して回転する。
【０１１１】
こうして、上記第１，第２のピニオン４１，４２と上記キャリヤ３８とが一体化することで両者の間には摩擦トルク等が生じない状態になり、上記第１のサンギヤ３９の入力トルクＴｉに対し上記キャリヤ３８の前輪側トルクＴＦ，上記第２のサンギヤ４０の後輪側トルクＴＢは、等トルク配分に歯車諸元が設定されていれば、この等トルク配分機能の歯車諸元による基準トルク配分，ＴＦ対ＴＢが略５０対５０のみに設定され、不等トルク配分に歯車諸元が設定されていれば、この不等トルク配分機能の歯車諸元による基準トルク配分に、ＴＦ対ＴＢが設定される。
【０１１２】
次に、前輪側回転数が後輪側回転数より大きくなるＮＦ＞ＮＢの旋回または前輪側スリップ時には、センターディファレンシャル装置３の上記第１，第２のピニオン４１，４２が遊星回転し、差動機能を有する歯車諸元により差動作用する。このため旋回時には、前後輪の回転数差が吸収されて、滑らかに旋回することになる。
【０１１３】
上記第１，第２のピニオン４１，４２の遊星回転に伴い、そのねじれ角の違いによるスラスト荷重が、上記第１，第２のピニオン４１，４２の一方の端面の部分に作用する。また、ギヤ噛合い点の分離，接線荷重の合成力が上記第１，第２のピニオン４１，４２，プラネタリピン４３の部分に作用して両者によりピニオン回転方向と反対の摩擦トルクと、これに基づく差動制限トルクが生じるようになる。
【０１１４】
そしてこの条件では、差動制限トルクがキャリア３８の回転を損うように作用することで、差動制限トルクが後輪側に移動して、トルク配分は基準トルク配分より後輪偏重になる。このため、旋回時の回頭性、操縦性が良くなり、また、直進時の前輪スリップ時にはスリップを防止するようになる。
【０１１５】
更に、後輪側回転数が前輪側輪回転数より大きいＮＢ＞ＮＦの後輪スリップ時には、センターディファレンシャル装置３の上記第１，第２のピニオン４１，４２が前後輪の回転数差により同様に遊星回転して摩擦トルクを発生する。
【０１１６】
ところでこの条件では、差動制限トルクがキャリヤ３８の回転を促すように作用して前輪側に移動するようになり、このため基準トルク配分より前輪側に多いトルク配分になって後輪スリップを防止する。
【０１１７】
ここで、上記遊星歯車機構による差動制限トルクは、入力トルクに対し比例的に生じるため、前後輪のトルクの大小に対して常に同じ割合になり、差動制限機能が常に一定の割合で発揮される。
【０１１８】
上述のようにセンターディファレンシャル装置３で分配された駆動力の一方の後輪側に分配された駆動力は、プロペラシャフト５，ドライブピニオン６を経て後輪左右駆動力配分部７に入力され、クラウンギヤ５２を通じてディファレンシャルケース５１に入力される。
【０１１９】
まず、差動制限機構部４４の機能について説明すると、上記ディファレンシャルケース５１に入力された駆動力は、第１のサンギヤ５６に入力され、第１，第２のピニオン５８，５９から第２のサンギヤ５７と、第１，第２のピニオン５８，５９を支持するキャリヤ５３とに分配されて伝達し、上記第２のサンギヤ５７の動力は、右ドライブ軸１７を介して右後輪１９に伝達される。また、上記キャリヤ５３の動力は左ドライブ軸１４を介して左後輪１６に伝達され駆動走行する。
【０１２０】
そこで、例えば左後輪回転数と右後輪回転数が等しいＮＬ＝ＮＲの直進走行では、上記第２のサンギヤ５７と上記キャリヤ５３とが同一方向に等速回転することで、上記第１，第２のピニオン５８，５９は遊星回転しなくなり一体化して回転する。
【０１２１】
こうして、上記第１，第２のピニオン５８，５９と上記キャリヤ５３とが一体化することで両者の間には摩擦トルク等が生じない状態になり、上記第１のサンギヤ５６の入力トルクに対し上記キャリヤ５３の左後輪トルク，上記第２のサンギヤ５７の右後輪トルクは、等トルク配分機能の歯車諸元による基準トルク配分が略５０対５０のみに設定される。
【０１２２】
次に、左後輪回転数が右後輪回転数より大きくなるＮＬ＞ＮＲの左後輪スリップ時には、上記第１，第２のピニオン５８，５９が遊星回転し、差動機能を有する歯車諸元により差動作用する。このため旋回時には、左右後輪の回転数差が吸収されて滑らかに旋回することになる。
【０１２３】
上記第１，第２のピニオン５８，５９の遊星回転に伴い、そのねじれ角の違いによるスラスト荷重が、上記第１，第２のピニオン５８，５９の一方の端面の部分に作用する。また、ギヤ噛合い点の分離，接線荷重の合成力が上記第１，第２のピニオン５８，５９，プラネタリピン６１の部分に作用して両者によりピニオン回転方向と反対の摩擦トルクと、これに基づく差動制限トルクが生じるようになる。
【０１２４】
そしてこの条件では、差動制限トルクがキャリア５３の回転を損うように作用することで、差動制限トルクが右後輪側に移動して、トルク配分は基準トルク配分より右後輪偏重になる。このため、直進時の左後輪スリップ時にはスリップを防止するようになる。
【０１２５】
更に、右後輪回転数が左後輪回転数より大きいＮＲ＞ＮＬの右後輪スリップ時には、上記第１，第２のピニオン５８，５９が左右後輪の回転数差により同様に遊星回転して摩擦トルクを発生する。
【０１２６】
ところでこの条件では、差動制限トルクがキャリヤ５３の回転を促すように作用して左後輪側に移動するようになり、このため基準トルク配分より左後輪に多いトルク配分になって右後輪スリップを防止する。
【０１２７】
ここで、上記遊星歯車機構による差動制限トルクは、入力トルクに対し比例的に生じるため、左右後輪のトルクの大小に対して常に同じ割合になり、差動制限機能が常に一定の割合で発揮される。
【０１２８】
次に、歯車機構部４５とクラッチ機構部４６について説明すると、上記ディファレンシャルケース５１が回転されると、このディファレンシャルケース５１に固定された第１の歯車軸５４が回転されて第１の歯車６２が回転され、この第１の歯車６１と噛合された第４の歯車６５が回転されて、第４の歯車軸６８とクラッチドラム６９が回転される。
【０１２９】
また、左輪出力側である上記キャリヤ５３の回転により、このキャリヤ５３に固定された第２の歯車軸５５が回転されて第２の歯車６３が回転され、この第２の歯車６３と噛合された第５の歯車６６が回転され、第５の歯車軸７０とクラッチハブ７１が回転される。
【０１３０】
さらに、上記右ドライブ軸１７の回転とともに第３の歯車６４が回転され、この第３の歯車６４と噛合された第６の歯車６７が回転され、第６の歯車軸７４，クラッチハブ７５が回転される。
【０１３１】
一方、後輪側油圧制御装置３４では、モータ８２によりオイルポンプ８３が駆動され、レギュレータ弁８４による作動油圧がデューティソレノイド弁８９とクラッチ制御弁８６とに導かれている。
【０１３２】
また、左右駆動力配分制御部３２では、、スロットル開度センサ２６，車速センサ２７，舵角センサ２９，前後加速度センサ３０，横加速度センサ３１、およびギヤ位置信号が入力処理され、前後の左右輪の最適な駆動力配分量が演算されている。
【０１３３】
そして、車両が右旋回状態で後輪側の左側に多く駆動力配分する場合、上記左右駆動力配分制御部３２から、上記後輪側油圧制御装置３４に対して、第１の油圧室７９側に演算した設定圧で油圧を加えるように信号が送られる。
【０１３４】
この結果、油圧管路３６ａを介して上記第１の油圧室７９に油圧が加えられ、第１のピストン７８が動作され、第１の油圧多板クラッチ７３が設定圧で連結され、上記ディファレンシャルケース５１からの駆動力は、上記第１の油圧多板クラッチ７３、第２の歯車６３を経て、上記キャリヤ５３、左ドライブ軸１４へと配分される。
【０１３５】
一方、車両が左旋回状態で後輪側の右側に多く駆動力配分する場合、上記左右駆動力配分制御部３２から、上記後輪側油圧制御装置３４に対して、第２の油圧室８１側に演算した設定圧で油圧を加えるように信号が送られる。
【０１３６】
この結果、油圧管路３６ｂを介して上記第２の油圧室８１に油圧が加えられ、第２のピストン８０が動作され、第２の油圧多板クラッチ７７が設定圧で連結され、上記ディファレンシ
ャルケース５１からの駆動力は、上記第２の油圧多板クラッチ７７、上記第３の歯車６４を経て、上記右ドライブ軸１７へと配分される。
【０１３７】
さらに、前記センターディファレンシャル装置３で分配された駆動力の他方の前輪側に分配された駆動力は、フロントドライブ軸１０，ドライブピニオン１１を経て前輪左右駆動力配分部１２に入力され、クラウンギヤ５２を通じてディファレンシャルケース５１に入力される。
【０１３８】
そして、上記前輪左右駆動力配分部１２も上記後輪左右駆動力配分部７と同様に、前輪側の差動機能を有して、上記左右駆動力配分制御部３２の信号が入力される前輪側の油圧制御装置３５からの制御油圧で、車両が右旋回状態で前輪側の左側に多く駆動力配分する場合には、油圧管路３７ａを介して上記第１の油圧室７９に油圧が加えられ、車両が左旋回状態で前輪側の右側に多く駆動力配分する場合には、油圧管路３７ｂを介して上記第２の油圧室８１に油圧が加えられる。
【０１３９】
以上のように本発明の実施の形態１によれば、左右駆動力配分装置は、左右輪方向の幅寸法の長大化が防止され、左右輪とアクスル軸間に配置される自在継手の交差角を小さくすることができ、耐久・信頼性を向上させることができる。
【０１４０】
また、左右駆動力配分装置は、構成部品点数も少なく、かつ従来のものを多く利用することができ、従来のものと互換性があり、コンパクトで、製造コストも低くすることができる。
【０１４１】
さらに、左右駆動力配分装置は、左右輪方向の幅寸法がコンパクトに構成されるため、サスペンションの構成部材や排気系部材との干渉や整備時に隙間が確保できるなど、車載性に優れる。
【０１４２】
また、左右駆動力配分装置は、ディファレンシャルケースから駆動力を両出力側にバイパスする構成になっているため、全体がコンパクトになる。
【０１４３】
さらに、左右駆動力配分装置は、３列の歯車列からなる駆動力配分機構を採用するため、ギヤ比の設定で駆動力配分が調整でき、車両の性格や狙いに容易に適合させることが可能である。
【０１４４】
また、左右駆動力配分装置の２組の油圧多板クラッチは、一体化された構造であり、軽量、コンパクトになる。
【０１４５】
また、差動制限機構部が歯車諸元（ねじれ角、圧力角等）やピニオンとキャリヤ、プラネタリピン等の回転摩擦面に作用する摩擦力でトルク感応型の差動制限機能を有することから、この差動制限機能と積極的な左右駆動力配分を兼ね備えた装置が実現される。
【０１４６】
また、センタディファレンシャル装置は、簡単な構造で部品点数も少なく、軽量コンパクトで、このため加工性、組立性に優れ、また動力伝達系の振動騒音に関しても有利になる。
【０１４７】
さらに、センタディファレンシャル装置と左右駆動力配分装置は共に軽量コンパクトであり、容易に一体にすることができ、軽量コンパクトな一体化ユニットが実現できる。
【０１４８】
また、センターディファレンシャル装置は、基準トルク配分を５０対５０の比率で前後輪に配分するように歯数を設定することができ、入力トルク比例式の差動制限トルクが前輪もしくは後輪へ走行状態や路面条件に応じて移動し、車両のスリップを防止して駆動力の確保や車両の尻振り等の挙動を防止し、走破性を向上させることができる。また、アクセル操作に対する車両の姿勢コントロールがしやすく、且つレスポンスも良くスポーティな走行を楽しむことができる。
【０１４９】
次に、図１２は本発明の実施の形態２による後輪左右駆動力配分部の拡大スケルトン図である。尚、本発明の実施の形態２は、前記発明の実施の形態１の後輪左右駆動力配分部のクラッチ機構部の位置を変更したものである。
【０１５０】
すなわち、図に示すように、第４の歯車６５の左側（ディファレンシャルケース５１側）に、上記第２の歯車６３に関する第１の油圧多板クラッチ７３を形成し、さらにその左側で上記第３の歯車６４に関する第２の油圧多板クラッチ７７を形成して２組一体のクラッチ機構部が構成されている。
【０１５１】
このように、クラッチ機構部の位置を変更することにより、左右駆動力配分部の左右方向長さをさらに短縮することが可能となる。
【０１５２】
次に、図１３は本発明の実施の形態３による後輪左右駆動力配分部の拡大スケルトン図である。尚、本発明の実施の形態３は、前記発明の実施の形態１の後輪左右駆動力配分部の歯車機構部の第１の歯車列のギヤ比を変更したものである。
【０１５３】
図に示す後輪左右駆動力配分部では、各歯車列のギヤ比が以下のように設定されている。、第１の歯車列は、ｚ4 ／ｚ1 ＝１／０．９、第２の歯車列は、ｚ5 ／ｚ2 ＝１、第３の歯車列は、ｚ6 ／ｚ3 ＝１で、第１の歯車列のギヤ比だけが減速比に設定されている。このため、各クラッチの連結による動作が、上記発明の実施の形態１の動作と逆に行われることになる。
【０１５４】
すなわち、車両が右旋回状態で後輪側の左側に多く駆動力配分する場合、上記左右駆動力配分制御部３２から、上記後輪側油圧制御装置３４に対して、第２の油圧室８１側に演算した設定圧で油圧を加えるように信号が送られる。
【０１５５】
この結果、油圧管路３６ｂを介して上記第２の油圧室８１に油圧が加えられ、第２のピストン８０が動作され、第２の油圧多板クラッチ７７が設定圧で連結され、上記ディファレンシャルケース５１からの駆動力は、右ドライブ軸１７側に少なくなり、上記キャリヤ５３側へ多く配分される。
【０１５６】
一方、車両が左旋回状態で後輪側の右側に多く駆動力配分する場合、上記左右駆動力配分制御部３２から、上記後輪側油圧制御装置３４に対して、第１の油圧室７９側に演算した設定圧で油圧を加えるように信号が送られる。
【０１５７】
この結果、油圧管路３６ａを介して上記第１の油圧室７９に油圧が加えられ、第１のピストン７８が動作され、第１の油圧多板クラッチ７３が設定圧で連結され、上記ディファレンシャルケース５１からの駆動力は、上記キャリヤ５３側に少なくなり、上記右ドライブ軸１７側へ多く配分される。
【０１５８】
次に、図１４は本発明の実施の形態４による後輪左右駆動力配分部の拡大スケルトン図である。尚、本発明の実施の形態４は、前記発明の実施の形態３の後輪左右駆動力配分部のクラッチ機構部の位置を変更したものである。
【０１５９】
すなわち、図に示すように、第４の歯車６５の左側（ディファレンシャルケース５１側）に、上記第２の歯車６３に関する第１の油圧多板クラッチ７３を形成し、さらにその左側で上記第３の歯車６４に関する第２の油圧多板クラッチ７７を形成して２組一体のクラッチ機構部が構成されている。
【０１６０】
このように、クラッチ機構部の位置を変更することにより、左右駆動力配分部の左右方向長さをさらに短縮することが可能となる。
【０１６１】
次に、図１５は本発明の実施の形態５による後輪左右駆動力配分部の拡大スケルトン図である。尚、本発明の実施の形態５は、前記発明の実施の形態２の歯車機構部を分離して形成するものである。
【０１６２】
すなわち、図に示すように、第２の歯車６３はキャリヤ５３に固定するのでははなく、左ドライブ軸１４に固定して構成される。このため、ディファレンシャルケース５１に対して、第１の歯車列と第３の歯車列が右側に、第２の歯車列が左側に分離して形成される。
【０１６３】
次に、図１６は本発明の実施の形態６による後輪左右駆動力配分部の拡大スケルトン図である。尚、本発明の実施の形態６は、前記発明の実施の形態４の歯車機構部を分離して形成するものである。
【０１６４】
すなわち、図に示すように、第２の歯車６３はキャリヤ５３に固定するのでははなく、左ドライブ軸１４に固定して構成される。このため、ディファレンシャルケース５１に対して、第１の歯車列と第３の歯車列が右側に、第２の歯車列が左側に分離して形成される。
【０１６５】
次に、図１７は本発明の実施の形態７による後輪左右駆動力配分部の拡大スケルトン図である。尚、本発明の実施の形態７は、駆動力配分を、一方の出力側に設けた２つの歯車列で行うようにしたものである。
【０１６６】
すなわち、図に示すように、ディファレンシャルケース５１には、第１の歯車６２が固定され、右ドライブ軸１７には、第２の歯車９５と第３の歯車９６とが固定され、これら第１，２，３の歯車はそれぞれ同一回転軸芯上の第４，５，６の歯車６５，９７，９８と噛合されている。
【０１６７】
上記第２の歯車９５は、上記第１の歯車６２より大径の歯車で、上記第３の歯車９６は上記第１の歯車６２より小径の歯車で形成されている。上記第１の歯車６２と上記第４の歯車６５で構成される第１の歯車列と、上記第２の歯車９５と上記第５の歯車９７で構成される第２の歯車列と、上記第３の歯車９６と上記第６の歯車９８で構成される第３の歯車列のギヤ比は、上記第１，２，３，４，５，６の歯車６２，９５，９６，６５，９７，９８の歯数をそれぞれｚ1 ，ｚ2 ，ｚ3 ，ｚ4 ，ｚ5 ，ｚ6 として、第１の歯車列は、ｚ4 ／ｚ1 ＝０．９、第２の歯車列は、ｚ5 ／ｚ2 ＝０．９×０．９、第３の歯車列は、ｚ6 ／ｚ3 ＝１に設定され、各ギヤ比を大きい順にならべると、１（第３の歯車列），０．９（第１の歯車列），０．９×０．９（第２の歯車列）で、ステップ比が０．９の一定になっている。尚、この値は他の値に設定しても良い。
【０１６８】
また、上記第１の歯車列と上記第２の歯車列間は、第１の油圧多板クラッチ９９で連結され動力伝達が可能に構成され、上記第１の歯車列と上記第３の歯車列間は、第２の油圧多板クラッチ１００で連結され動力伝達が可能に構成されている。
【０１６９】
上記２組一体の油圧多板クラッチの容量は、上記ギヤ比に応じて予め不等値に設定しても良く、一方の多板クラッチを小型にすることもできる。
【０１７０】
このように構成することにより、車両が右旋回状態で後輪側の左側に多く駆動力配分する場合、上記左右駆動力配分制御部３２から、上記後輪側油圧制御装置３４に対して、第１の油圧室７９側に演算した設定圧で油圧を加えるように信号が送られる。
【０１７１】
この結果、油圧管路３６ａを介して上記第１の油圧室７９に油圧が加えられ、第１のピストン７８が動作され、第１の油圧多板クラッチ７３が設定圧で連結され、上記ディファレンシャルケース５１からの駆動力は、右ドライブ軸１７側に少なくなり、上記キャリヤ５３側へ多く配分される。
【０１７２】
一方、車両が左旋回状態で後輪側の右側に多く駆動力配分する場合、上記左右駆動力配分制御部３２から、上記後輪側油圧制御装置３４に対して、第２の油圧室８１側に演算した設定圧で油圧を加えるように信号が送られる。
【０１７３】
この結果、油圧管路３６ｂを介して上記第２の油圧室８１に油圧が加えられ、第２のピストン８０が動作され、第２の油圧多板クラッチ７７が設定圧で連結され、上記ディファレンシャルケース５１からの駆動力は、上記右ドライブ軸１７側へ多く配分される。
【０１７４】
本発明の実施の形態７によっても、左右駆動力配分装置は、３列の歯車列のギヤ比を所定の関係に設定することで、２組の油圧多板クラッチのドライブプレートとドリブンプレートの相対回転差を同一にすることがてきるため、同じ摩擦特性（速度と動摩擦係数の関係）が得られる範囲を利用でき、制御精度を高くすることができる。
【０１７５】
以上の各発明の実施の形態に示すように、歯車機構部を構成する各歯車列の位置、クラッチ機構部の位置は様々に設定され、上記各発明の実施の形態に示す以外の位置であっても良い。また、上記各発明の実施の形態で説明する歯車機構部の各歯車列のギヤ比は、他の値に設定するものでも良い。
【０１７６】
上記各発明の実施の形態で説明した以外の車両、ＦＦ（フロントエンジン・フロントドライブ）車，ＲＲ（リヤエンジン・リヤドライブ）車，４ＷＤ車の前後輪のどちらかの左右駆動力配分を行う車両等においても本発明は適用できる。
【０１７７】
また、４ＷＤ車に適用する場合のセンターディファレンシャル装置は、上記説明した以外のものであっても良い。
【０１７８】
【発明の効果】
以上説明したように本発明によれば、左右の出力軸方向の長大化を防止し、左右輪とアクスル軸間に配置される自在継手の交差角を小さくすることができ、サスペンションの構成部材や排気系部材との干渉や整備時に隙間が確保できるなど車載性に優れ、構成部品点数も少なく小型・軽量で、従来の差動装置と装着互換性を有し、製造コスト上有利であり、また、差動制限機能も有し、バイパストルクの調整・設定も容易で制御精度が高く耐久・信頼性に優れるという効果を奏する。
【図面の簡単な説明】
【図１】本発明の実施の形態１による４ＷＤ車の全体の概略構成を示す説明図
【図２】本発明の実施の形態１による後輪左右駆動力配分部の拡大断面図
【図３】本発明の実施の形態１によるセンターディファレンシャル装置の差動機能説明のための各部の概略図
【図４】本発明の実施の形態１による第１のサンギヤを固定した際の動作説明図
【図５】本発明の実施の形態１による第２のサンギヤを固定した際の動作説明図
【図６】本発明の実施の形態１によるセンターディファレンシャル装置の動力分配機能、差動制限機能説明のための各部の概略図
【図７】本発明の実施の形態１による各ギヤにより生じる荷重の説明図
【図８】本発明の実施の形態１による後輪側回転数よりも前輪側回転数の方が大きい場合の説明
【図９】本発明の実施の形態１による後輪側回転数よりも前輪側回転数の方が小さい場合の説明図
【図１０】本発明の実施の形態１による左右駆動力配分の油圧制御装置の構成説明図
【図１１】本発明の実施の形態１による左右駆動力配分制御部の機能ブロック説明図
【図１２】本発明の実施の形態２による後輪左右駆動力配分部の拡大スケルトン図
【図１３】本発明の実施の形態３による後輪左右駆動力配分部の拡大スケルトン図
【図１４】本発明の実施の形態４による後輪左右駆動力配分部の拡大スケルトン図
【図１５】本発明の実施の形態５による後輪左右駆動力配分部の拡大スケルトン図
【図１６】本発明の実施の形態６による後輪左右駆動力配分部の拡大スケルトン図
【図１７】本発明の実施の形態７による後輪左右駆動力配分部の拡大スケルトン図
【符号の説明】
１ エンジン
２ 自動変速装置
３ センターディファレンシャル装置
４ リヤドライブ軸
５ プロペラシャフト
６ ドライブピニオン
７ 後輪左右駆動力配分部
８ トランスファドライブギヤ
９ トランスファドリブンギヤ
１０ フロントドライブ軸
１１ ドライブピニオン
１２ 前輪左右駆動力配分部
１４ 左ドライブ軸
１６ 左後輪
１７ 右ドライブ軸
１９ 右後輪
２０ 左ドライブ軸
２１ 左前輪
２２ 右ドライブ軸
２３ 右前輪
２６ スロットル開度センサ
２７ 車速センサ
２９ 舵角センサ
３０ 前後加速度センサ
３１ 横加速度センサ
３２ 左右駆動力配分制御部
３４ 後輪側油圧制御装置
３５ 前輪側油圧制御装置
４４ 差動制限機構部
４５ 歯車機構部
４６ クラッチ機構部
５１ ディファレンシャルケース
５２ クラウンギヤ
５３ キャリヤ
５４ 第１の歯車軸
５５ 第２の歯車軸
５６ 第１のサンギヤ
５７ 第２のサンギヤ
５８ 第１のピニオン
５９ 第２のピニオン
６０ ピニオン部材
６１ プラネタリピン
６２ 第１の歯車
６３ 第２の歯車
６４ 第３の歯車
６５ 第４の歯車
６６ 第５の歯車
６７ 第６の歯車
６８ 第４の歯車軸
６９ クラッチドラム
７０ 第５の歯車軸
７１ クラッチハブ
７２ プレート
７３ 第１の油圧多板クラッチ
７４ 第６の歯車軸
７５ クラッチハブ
７６ プレート
７７ 第２の油圧多板クラッチ
７８ 第１のピストン
７９ 第１の油圧室
８０ 第２のピストン
８１ 第２の油圧室[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a differential device for a four-wheel drive vehicle or a two-wheel drive vehicle, and more particularly to a vehicle left / right driving force distribution device capable of variably controlling a driving force distribution ratio to left and right wheels.
[0002]
[Prior art]
In recent years, various types of differential limiting devices have been developed to ensure driving force and improve running stability and motor performance when crossing rough roads, steep slopes, and traveling on split μ roads. It has been put into practical use. Furthermore, recently, a technique has been proposed in which the torque distribution of the left and right wheels is positively adjusted to improve the turning performance of the vehicle.
[0003]
For example, in Japanese Patent Application Laid-Open No. 5-77653, a ring gear is formed on the inner periphery of a differential case, a sun gear is attached to a second output shaft, and a carrier that supports a planetary gear is attached to the first output shaft. A differential device formed of a double pinion type planetary gear mechanism is provided with a driving force transmission control mechanism for adjusting the left and right driving force distribution, and the driving force transmission control mechanism is A transmission mechanism that is attached to the output shaft and changes the rotational speed of the left and right output shafts; a driving force transmission auxiliary member that is shifted by the transmission mechanism and connected to rotate at a different speed from the left and right output shafts; A multi-plate clutch mechanism that adjusts left and right driving force distribution is provided, and the multi-plate clutch mechanism and the differential mechanism are disposed in the same casing. The multi-plate clutch mechanism is separated from the clutch portion and the piston portion, and the clutch portion is disposed in the differential case and the piston portion is disposed outside the case.
[0004]
Japanese Patent Laid-Open No. 5-345535 discloses a double pinion type planetary gear mechanism differential mechanism in which a driving force transmission control mechanism for left and right wheels is interposed between rotating shafts for left and right wheels. The speed increasing mechanism that increases the speed of one of the rotating shafts and outputs it to the first intermediate shaft is integrated with the speed reducing mechanism that decelerates one of the rotating shafts and outputs it to the second intermediate shaft. In the figure, the speed increasing / decreasing mechanism and the first and second transmission torque capacity varying means are integrated, and the first and second transmission torque capacity varying means are integrated adjacent to each other. The transmission torque capacity varying means is constituted by an electronically controlled hydraulic multi-plate clutch.
[0005]
Further, in Japanese Patent Application Laid-Open No. 1-182127, a pair of pinions (bevel gears) are rotatably provided on a pinion shaft fixed in a differential case rotated by an input shaft, and left and right side gears (bevel gears) are provided on the pinion. ) To form a differential, and a hydraulic multi-plate clutch is provided on each of the left wheel side and the right wheel side of the intermediate shaft that rotates together with the differential case, and the left wheel side output shaft, the intermediate shaft, and the right wheel side output shaft. And the intermediate shaft are connected via respective hydraulic multi-plate clutches, and the torque transmission amount of the left and right wheels is variably controlled in accordance with the motion state of the vehicle.
[0006]
[Problems to be solved by the invention]
However, in the technique disclosed in Japanese Patent Laid-Open No. 5-77653, the drive plate and the driven plate of the hydraulic multi-plate clutch are alternately stacked on both sides of the differential case formed on the inner diameter portion of the hypoid ring gear, and a double pinion is provided at the center portion thereof. Drive differential transmission member and a multi-plate clutch that adjusts the left and right drive force distribution to rotate at a different speed from the output shafts for the left and right wheels on the outside of the differential case Since the mechanisms are arranged in series, there is a problem that the dimensions in the left and right wheel directions become long.
[0007]
And since the output shafts of the left and right wheels become longer in the wheel direction, the overall length of the drive shaft with constant velocity joints on both sides is shortened, the difference between the position of the axle and the output shaft determined from the vehicle body layout, and the vertical movement according to the suspension stroke The joint angle (bending angle) of the drive shaft increases due to dynamic changes such as vertical shifts according to changes in stroke and vehicle type, vehicle rebound, etc., resulting in a decrease in drive shaft strength, transmission efficiency, and joints. May cause vibration and noise problems.
[0008]
Also, because the planetary gear type driving force transmission auxiliary members for distributing left and right wheel driving force are arranged on both sides of the differential case, the entire device becomes larger, the structure becomes complicated, the number of components increases, and the manufacturing cost and mass increase. From the viewpoint of
[0009]
Furthermore, there is a restriction on the outer diameter of the clutch disk of the hydraulic multi-plate clutch. To increase the clutch torque capacity, either increase the number of clutches significantly or increase the pressure receiving area of the hydraulic piston. It is necessary to take measures. For this reason, since it is necessary to increase the torque capacity in a high-power vehicle, the entire apparatus is undesirably increased in size, mass, and cost.
[0010]
In addition, since the structure in the left-right direction is separated into three blocks: a left planetary gear and a left piston, a differential gear and a hydraulic multi-plate clutch in a differential case, and a right planetary gear and a right piston, the lubrication balance and lubrication method are There is a problem such as difficulty.
[0011]
Further, in the technique disclosed in Japanese Patent Laid-Open No. 5-345535, a compound planetary gear in which a double pinion type differential device, a speed increasing mechanism and a speed reducing mechanism are integrated in the left and right output shaft directions, and a compound planetary gear integrated with this compound planetary gear. Two sun gears are connected to two sets of hydraulic multi-plate clutches so that power can be transmitted, and two sets of hydraulic multi-plate clutches are integrated adjacent to each other. Since the compound planetary gear type speed increasing / decreasing mechanism, two sets of hydraulic multi-plate clutches, and the like are configured in series, the width dimension of the left and right driving force distribution device becomes long, and there is a problem similar to the above-described prior art.
[0012]
In addition, the driving force to bypass is a system that moves between the left wheel output shaft and the right wheel output shaft of the differential gear via an increase / decrease mechanism and a hydraulic multi-plate clutch, effectively making the transmission torque capacity of the multi-plate clutch effective. Available. However, since the double pinion type differential device is composed of a ring gear having a large diameter, a large and small pinion, and a sun gear, a differential case (differential on one side) is restricted due to space limitations when stored in the inner diameter part of the ring gear of the hypoid gear. A hypoid ring gear is bolted with three pieces to a ring gear which is an input element of a gear and a bearing support member is formed on the other. For this reason, the structure is supported by a bearing directly below the hypoid ring gear, and the balance between the left and right bearing capacities is deteriorated, which is disadvantageous in terms of durability, reliability, gear noise, and the like.
[0013]
Bearings at both ends of the hypoid ring gear in the direction of the left and right output shafts, bearings for the input sun gear of the compound planetary gear, the support for the triple pinion, the walls of the hydraulic chamber in which the hydraulic pistons of the two sets of hydraulic multi-plate clutch are housed, both ends It is necessary to provide a bearing that supports the output shaft and stationary member such as an oil seal, and at least five rigid walls must be formed. The overall structure is complicated, and the manufacturing cost, mass increase, and lubrication balance are required. It is disadvantageous in terms of stirring loss due to the clutch during turning. In addition, the divided structure of the case of the entire differential device becomes complicated, the sealing surface of the lubricating oil is increased, and there are problems such as reliability, high manufacturing costs, and poor compatibility with current production parts.
[0014]
Further, in the technique disclosed in Japanese Patent Laid-Open No. 1-182127, the driving force is distributed through hydraulic multi-plate clutches provided independently on the left and right wheels, respectively. There is a problem that a clutch has to be provided, and the lateral dimension of the vehicle becomes long and the number of parts increases.
[0015]
The present invention has been made in view of the above circumstances, can prevent an increase in the length of the left and right output shaft directions, can reduce the intersection angle of the universal joint disposed between the left and right wheels and the axle shaft, Excellent in-vehicle performance, such as interference with components and exhaust system members and ensuring clearance during maintenance, small number of components, light weight, mounting compatibility with conventional differentials, and advantageous in manufacturing cost Another object of the present invention is to provide a vehicle left / right driving force distribution device that has a differential limiting function, that is easy to adjust and set a bypass torque, has high control accuracy, and is excellent in durability and reliability.
[0016]
[Means for Solving the Problems]
In order to achieve the above object, a left / right driving force distribution device for a vehicle according to the present invention as claimed in claim 1 forms a first gear train by meshing a first sun gear on the driving force input side with a first pinion. The second sun gear on the output side of either the left wheel side or the right wheel side meshes with a second pinion integral with the first pinion to form a second gear train, and the first and first gears The two pinions are pivotally supported by the carrier on the other output side, and the separation load acting on the gear meshing point of the first gear train and the second gear train is applied to the shafts of the first and second pinions. A condition in which there is a differential limiting device that generates a differential limiting torque proportional to the input torque between the left and right wheels with frictional force obtained by acting on the support portion, and there is no differential between the left and right wheels on the input side of the driving force A first gear for generating a reference rotational speed for the first rotating member is provided, and the one output side A second gear that rotates the second rotating member at a second rotational speed different from the reference rotational speed on the condition that there is no differential between the left and right wheels on either of the other output sides is provided. A third gear for rotating the third rotating member at a third rotational speed different from the reference rotational speed on the condition that there is no differential between the left and right wheels on either the output side or the other output side; The first rotating member, the second rotating member, and the third rotating member are disposed on the same rotating shaft parallel to the rotating shafts on both output sides, and the first rotating member And the second rotating member has a variable transmission capacity clutch mechanism, and the first rotating member and the third rotating member have a variable transmission capacity clutch mechanism. It is integrally formed on the top.
[0017]
According to a second aspect of the present invention, there is provided the vehicle left / right driving force distribution device according to the first aspect, wherein the first gear train and the second gear train are engaged with each other. The frictional force obtained by acting the difference in thrust load acting on the first point on one end face of the first and second pinions, and the gear meshing point of the first gear train and the second gear train Difference in generating differential limiting torque proportional to the input torque between the left and right wheels by the frictional force obtained by applying the combined force of the acting separation load and tangential load to the shaft support portions of the first and second pinions. It is equipped with a movement limiting device.
[0018]
Furthermore, the vehicle left / right driving force distribution device according to the present invention as defined in claim 3 is the vehicle left / right driving force distribution device according to claim 1 or 2, wherein the two sets of clutch mechanisms are hydraulic multi-plate clutches. It is formed by at least one of an electromagnetic clutch and a variable transmission capacity type coupling.
[0019]
According to a fourth aspect of the present invention, there is provided the vehicle left / right driving force distribution device according to any one of the first, second, and third aspects, wherein the two clutch mechanisms are used. The transmission capacity is variably set according to the running state of the vehicle and the road surface condition.
[0020]
Furthermore, the left and right driving force distribution device for a vehicle according to the present invention as set forth in claim 5 is the vehicle left and right driving force distribution device according to any one of claims 1, 2, 3, and 4, wherein the second gear Is provided on the one output side, the third gear is provided on the other output side, and the second rotation speed of the second rotation member and the second rotation member under the condition that there is no differential between the left and right wheels. The third rotation speed of the third rotating member is set to the same value.
[0021]
A vehicle left / right driving force distribution device according to the present invention as set forth in claim 6 is the vehicle left / right driving force distribution device according to any one of claims 1, 2, 3 and 4, wherein the second gear And the third gear are provided on either the one output side or the other output side, and the second rotation speed of the second rotation member under the condition that there is no differential between the left and right wheels. And the third rotational speed of the third rotating member is obtained by setting the rotational speed on one side to a value larger than the reference rotational speed and the rotational speed on the other side smaller than the reference rotational speed. is there.
[0022]
Further, the vehicle left-right driving force distribution device according to the present invention as set forth in claim 7 is the vehicle left-right driving force distribution device according to claim 6, wherein the third gear under the condition that there is no differential between the left and right wheels. The ratio of the rotation speed of the third rotation member and the third rotation speed of the third rotation member, the ratio of the rotation speed of the first gear and the reference rotation speed of the first rotation member, and the rotation of the second gear The step ratio between the speed and the ratio of the second rotational speed of the second rotating member is set constant.
[0023]
According to the eighth aspect of the present invention, there is provided the vehicle left-right driving force distribution device according to the seventh aspect, wherein the two sets of clutch mechanisms are provided with a predetermined transmission capacity difference. Formed.
[0024]
In the left and right driving force distribution device for a vehicle according to claim 1, when driving force is input to the differential limiting device, power is distributed to one of the output side of the left wheel side and the right wheel side and the other output side. For example, the one output side is set as the left wheel side, and the other output side is set as the right wheel side. In the case where the friction coefficient of the road surface where the left wheel and the right wheel contact each other is greatly different between the left and right, when the left wheel and the right wheel are differentially rotated, the gear meshing of the first gear train and the second gear train is performed. The frictional force generated by the separation load acting on the point acting on the shaft support portions of the first and second pinions increases in proportion to the input torque, and this frictional force causes the first and second pinions to Differential limiting torque is generated in the opposite direction of rotation. This differential limiting torque is torque-distributed so as to move to the right wheel side when the left wheel slips and to the left wheel side when the rotation speed of the right wheel is large so as to prevent slipping. Further, the first gear provided on the input side is rotated by the driving force, and the first rotating member is rotated at the reference rotation speed. Further, the second rotating member is rotated by the rotation of the second gear provided on either the one output side or the other output side, and either the one output side or the other output side is rotated. The third rotating member is rotated by the rotation of the provided third gear. Each of these rotating members is rotated on the same rotation axis parallel to the rotation axis on both output sides, and each rotation speed is a value set under the condition that there is no differential between the left and right wheels. Then, the first rotating member and the second rotating member are connected in a variable transmission capacity by one clutch mechanism of two sets of clutch mechanisms formed integrally on the rotating shaft core of each rotating member. Alternatively, the torque distribution of the left and right wheels is performed by connecting the first rotating member and the third rotating member in a variable transmission capacity with the other clutch mechanism.
[0025]
Further, the vehicle left / right driving force distribution device according to claim 2 is the vehicle left / right driving force distribution device according to claim 1, wherein the differential limiting device is a gear of the first gear train and the second gear train. The frictional force obtained by applying the difference in thrust load acting on the meshing point to one end surface of the first and second pinions, and the gear meshing point of the first gear train and the second gear train A differential limiting torque proportional to the input torque is generated between the left and right wheels by the frictional force obtained by applying the combined force of the acting separation load and tangential load to the shaft support portions of the first and second pinions.
[0026]
Further, the vehicle left / right driving force distribution device according to claim 3 is the vehicle left / right driving force distribution device according to claim 1 or 2, wherein the two sets of clutch mechanisms are a hydraulic multi-plate clutch and an electromagnetic clutch. It is formed by at least one of a transmission capacity variable type coupling to obtain a clutch function.
[0027]
Further, the vehicle left-right driving force distribution device according to claim 4 is the vehicle left-right driving force distribution device according to any one of claims 1, 2, 3, wherein the transmission capacity of the two sets of clutch mechanisms is It is variably set according to the running state of the vehicle and the road surface condition.
[0028]
Further, the vehicle left / right driving force distribution device according to claim 5 is the vehicle left / right driving force distribution device according to any one of claims 1, 2, 3 and 4, wherein the second gear is moved to the one side. And the third gear is provided on the other output side, and the second rotational speed of the second rotating member and the third speed on the condition that there is no differential between the left and right wheels. Set the third rotational speed of the rotating member to the same value and connect the rotating member on the output side to the left wheel with the clutch mechanism, or connect the rotating member on the output side to the right wheel with the clutch mechanism To distribute torque.
[0029]
The vehicle left / right driving force distribution device according to claim 6 is the vehicle left / right driving force distribution device according to any one of claims 1, 2, 3 and 4, wherein the second gear and the first gear are arranged. 3 gears are provided on either the one output side or the other output side, and the second rotation speed of the second rotation member and the second rotation speed under the condition that there is no differential between the left and right wheels. The third rotation speed of the third rotation member is set such that the rotation speed on one side is set to a value larger than the reference rotation speed and the rotation speed on the other side is set to a value smaller than the reference rotation speed. By connecting the first rotating member and the second rotating member on the side having both the third gear and the third gear by the clutch mechanism, or by connecting the first rotating member and the third rotating member by the clutch mechanism Distributes torque between left and right wheels.
[0030]
Furthermore, the vehicle left-right driving force distribution device according to claim 7 is the vehicle left-right driving force distribution device according to claim 6, wherein the rotational speed of the third gear under the condition that there is no differential between the left and right wheels. And the ratio of the third rotational speed of the third rotating member, the ratio of the rotational speed of the first gear and the reference rotational speed of the first rotating member, the rotational speed of the second gear, and the above The step ratio between the second rotational speed and the second rotational speed ratio of the second rotating member is set to be constant and the torque distribution is controlled reliably and easily.
[0031]
The vehicle left / right driving force distribution device according to claim 8 is the vehicle left / right driving force distribution device according to claim 7, wherein the two sets of clutch mechanisms provide a predetermined transmission capacity difference.
[0032]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
1 to 11 show Embodiment 1 of the present invention, FIG. 1 is an explanatory diagram showing the overall schematic configuration of a 4WD vehicle, FIG. 2 is an enlarged sectional view of a rear wheel left / right driving force distribution unit, and FIG. FIG. 4 is an operation explanatory diagram when the first sun gear is fixed, FIG. 5 is an operation explanatory diagram when the second sun gear is fixed, and FIG. 6 is an operation explanatory diagram when the second sun gear is fixed. Schematic diagram of each part for explaining the power distribution function and differential limiting function of the center differential device, FIG. 7 is an explanatory diagram of the load generated by each gear, and FIG. 8 is the front wheel side rotational speed rather than the rear wheel side rotational speed. FIG. 9 is an explanatory diagram when the front wheel side rotational speed is smaller than the rear wheel side rotational speed, FIG. 10 is a structural explanatory diagram of a hydraulic control device for right and left driving force distribution, and FIG. It is a functional block explanatory drawing of a force distribution control part. In the first embodiment of the present invention, left and right driving force distribution devices are provided on the front and rear wheels of a 4WD (four-wheel drive) vehicle, and the center differential device is composed of a composite planetary gear type.
[0033]
In FIG. 1, reference numeral 1 denotes an engine disposed at the front of the vehicle, and the driving force of the engine 1 is transmitted from an automatic transmission device (including a torque converter and the like) 2 behind the engine 1 through a transmission output shaft 2a. It is transmitted to the center differential device 3, and is input from the center differential device 3 to the rear wheel left / right driving force distribution unit 7 via the rear drive shaft 4, the propeller shaft 5, and the drive pinion 6. A gear 8, a transfer driven gear 9, and the transfer driven gear 9 and the drive pinion 11 provided on the front drive shaft 10 together with the transfer driven gear 9 are configured to be input to the front wheel left / right driving force distribution unit 12. Here, the automatic transmission 2, the center differential device 3, the front wheel left / right driving force distribution unit 12 and the like are integrally provided in the case 13.
[0034]
The driving force input to the rear wheel left / right driving force distribution unit 7 passes through the left drive shaft 14 and the left axle shaft 15 to the left rear wheel 16, and passes through the right drive shaft 17 and the right axle shaft 18 to the right rear wheel 19. It is to be transmitted.
[0035]
The driving force input to the front wheel left / right driving force distribution unit 12 is transmitted to the left front wheel 21 via the left drive shaft 20 and to the right front wheel 23 via the right drive shaft 22.
[0036]
A throttle body 25 communicating with the intake manifold 24 of the engine 1 is provided with a throttle opening sensor 26 for detecting the opening (throttle opening) of a throttle valve (not shown), and the case 13 has a rear wheel output. A vehicle speed sensor 27 for detecting the shaft rotational speed as the vehicle speed is provided.
[0037]
The steering column of the steering wheel 28 is provided with a steering angle sensor 29 that detects the steering angle θ, and further includes a longitudinal acceleration sensor 30 that detects the longitudinal acceleration of the vehicle, and a lateral acceleration sensor that detects the lateral acceleration. 31 is provided.
[0038]
Each of the sensors is connected to a left and right driving force distribution control unit 32 that controls the driving force distribution of the front and rear left and right wheels. A gear position signal is taken out from a control unit 33 such as a TCU (transmission control unit) or an ECU (engine control unit) and is input to the left / right driving force distribution control unit 32.
[0039]
The left and right driving force distribution control unit 32 determines the running state and road surface condition of the vehicle based on each input signal, and obtains the optimal left and right wheel driving force distribution amount performed by the rear wheel left and right driving force distribution unit 7. A signal is output to the hydraulic control device 34 on the side, while an optimal left / right wheel driving force distribution amount performed by the front wheel left / right driving force distribution unit 12 is obtained and a signal is output to the hydraulic control device 35 on the front wheel side. ing.
[0040]
The rear wheel side hydraulic control device 34 and the front wheel side hydraulic control device 35 have substantially the same structure, and each receives a signal from the left and right driving force distribution control unit 32, and the rear wheel side hydraulic control device 34 receives the rear wheel. The left and right driving force distribution unit 7 is configured to apply hydraulic pressure through a pair of hydraulic lines 36, and the front wheel side hydraulic control device 35 is connected to the front wheel left and right driving force distribution unit 12 through a pair of hydraulic lines 37. It is configured to apply hydraulic pressure.
[0041]
The center differential device 3 is provided in the rear of the case 13, and the transmission output shaft 2a is rotatably inserted from the front of the carrier 38 accommodated rotatably, while the rear drive shaft 4 is inserted from the rear. Is inserted freely.
[0042]
A large-diameter first sun gear 39 is formed at the rear end portion of the transmission output shaft 2a on the input side, and a small-diameter second gear is formed at the front end portion of the rear drive shaft 4 that outputs to the rear wheels. A sun gear 40 is formed, and the first sun gear 39 and the second sun gear 40 are stored in the carrier 38.
[0043]
The first sun gear 39 meshes with the first pinion 41 having a small diameter to form a first gear train, and the second sun gear 40 meshes with the second pinion 42 having a large diameter. The gear train is formed.
[0044]
The first pinion 41 and the second pinion 42 are integrally formed, and a plurality of pairs (for example, three pairs) of the pinions are rotatably supported by respective planetary pins 43 fixed to the carrier 38. ing.
[0045]
Further, the carrier 38 is configured such that the transfer drive gear 8 is connected to the front end, and outputs from the carrier 38 to the front wheels.
[0046]
That is, the driving force from the transmission output shaft 2a is transmitted to the first sun gear 39 and is output from the second sun gear 40 to the rear drive shaft 4, and from the carrier 38 to the transfer drive gear 8 and the transfer driven gear. 9 is a composite planetary gear type center differential device that outputs to the front drive shaft 10 through 9.
[0047]
The composite planetary gear type center differential device has the number of teeth of the first and second sun gears 39 and 40 and a plurality of the first and second pinions 41 and 42 arranged around the sun gears 39 and 40. It has a differential function by setting appropriately.
[0048]
Further, by appropriately setting the meshing pitch circle radius between the first and second sun gears 39 and 40 and the first and second pinions 41 and 42, the reference torque distribution is an equal torque of 50:50 front and rear. It has the function of unequal torque distribution that is biased to either distribution or front and back.
[0049]
Further, the first and second sun gears 39 and 40 and the first and second pinions 41 and 42 are, for example, helical gears, and the twist angles of the first gear train and the second gear train are set. The thrust load remains without canceling out the thrust load and the friction torque between the pinion end faces is separated by the meshing between the first and second pinions 41 and 42 and the surface of the planetary pin 43, and the tangential load. A differential limiting function is provided by setting so that the resultant force acts and frictional torque is generated to obtain a differential limiting torque proportional to the input torque.
[0050]
Next, the differential function of the center differential device 3 will be described with reference to FIGS. 3, 4, and 5.
First, when the first sun gear 39 is fixed as shown in FIG. 4, on the circumference of the radius rs2,
(Arc KF) = (Arc CF) − (Arc CK) (1)
When the second sun gear 40 is fixed as shown in FIG. 5, on the circumference of the radius rs1,
(Arc arc ZF) = (Arc arc BF) − (Arc arc BZ) (2)
Is established.
[0051]
Here, the angular velocities ωs1, ωs2 of the first and second sun gears 39, 40, the meshing pitch circle radii rs1, rs2, the angular velocities ωp1, ωp2 of the first and second pinions 41, 42, and the meshing pitch circular radii rp1, rp2 Assuming that the angular velocity ωc of the carrier 38, the equation (1) is
ωs 2 · rs 2 = −ωp 2 · rp 2 + ωc · rs 2 (3)
And (2) is
ωs 1 · rs 1 = −ωp 1 · rp 1 + ωc · rs 1 (4)
become.
[0052]
Therefore, since the first and second pinions 41 and 42 are integrated and ωp1 = ωp2, the above equations (3) and (4) are rearranged.
ωc · (rs 2 −rs 1 · rp 2 / rp 1)
= Ωs 2 · rs 2 −ωs 1 · rs 1 · rp 2 / rp 1 (5)
Is established.
[0053]
Here, the angular speed ωs1 of the first sun gear 39 is the input rotational speed Ni, the angular speed ωc of the carrier 38 is the front wheel side rotational speed NF, the angular speed ωs2 of the second sun gear 40 is the rear wheel side rotational speed NB, and the first and second speeds. If the meshing pitch circle radii rs1 and rs2 of the sun gears 39 and 40 and the meshing pitch circle radii rp1 and rp2 of the first and second pinions 41 and 42 are replaced with the number of teeth Zs1, Zs2, Zp1, and Zp2, respectively (5 )ceremony,
NF ・ (Zs2−Zs1 ・ Zp2 / Zp1)
= NB.Zs2-Ni.Zs1.Zp2 / Zp1 (5) '
It becomes.
[0054]
If the number of teeth is Zp1 = 24, Zp2 = 24, Zs1 = 30, Zs2 = 15,
NB + NF = 2Ni
When Ni ≠ 0, NB>Ni> NF or NF>Ni> NB is established, and the front wheel side rotational speed NF and the rear wheel side rotational speed NB both have the same rotational direction and the differential. To establish.
[0055]
Next, the equal torque distribution function will be described with reference to FIG. 6, FIG. 7, FIG. 8, and FIG.
The input torque of the first sun gear 39 is Ti, the meshing pitch circle radius is rs1, the front wheel side torque of the carrier 38 is TF, the meshing pitch circle radii of the first and second pinions 41, 42 are rp1, rp2, When the rear wheel side torque of the sun gear 40 is TB and its meshing pitch circle radius is rs2,
Ti = TF + TB (6)
rs1 + rp1 = rs2 + rp2 (7)
Is established.
[0056]
The tangential load P acting on the meshing point between the first sun gear 39 and the first pinion 41 is the mesh between the tangential load P1 acting on the carrier 38 and the second sun gear 40 and the second pinion 42. Equal to the sum of the tangential loads P2 acting on the points.
P = Ti / rs1
P1 = TF / (rs1 + rp1)
P2 = TB / rs2
Ti / rs1 = {(TF / (rs1 + rp1)} + TB / rs2 (8)
Substituting the above equations (6) and (7) into the above equation (8) and rearranging,
TF = (1−rp 1 · rs 2 / rs 1 · rp 2) · Ti (9)
TB = (rp1 · rs2 / rs1 · rp2) · Ti (10)
It becomes. Therefore, the reference torque distribution of the front wheel side torque TF and the rear wheel side torque TB can be freely performed by the meshing pitch circle radius of the first and second sun gears 39, 40 and the first and second pinions 41, 42. It can be seen that it can be set.
[0057]
The meshing pitch circle radii rs1, rs2, rp1, and rp2 are replaced with the number of teeth Zs1, Zs2, Zp1, and Zp2 of each gear, and the number of teeth is substituted for each number of teeth (Zp1 = 24, Zp2 = 24, Zs1 = 30, Zs2 = 15), TF = 0.5 · Ti
TB = 0.5 · Ti
become. Therefore, the front and rear wheel torque distribution is approximately 50 to 50, and the reference torque distribution can be sufficiently set to the equal torque distribution.
[0058]
Further, the differential limiting function will be described. The first and second sun gears 39 and 40 and the first and second pinions 41 and 42 are helical gears having a predetermined twist angle. The planetary pin 43 is made without canceling the thrust loads acting on the meshing points with the first and second sun gears 39 and 40 by making the torsion angles of the first and second pinions 41 and 42 different. A sliding frictional force is generated at both end surfaces. Further, the combined force of the separation load and the tangential load acting on the meshing points of the first gear train and the second gear train is applied to the first and second pinions 41 and 42 and the planetary pin 43. As a result, rolling friction force is generated. These frictional forces generate a friction torque proportional to the input torque, that is, a differential limiting torque in the opposite direction to the pinion rotation.
[0059]
Here, the pinion rotation direction changes depending on the magnitude relationship between the front wheel side rotation speed NF and the rear wheel side rotation speed NB, and accordingly, the degree of application of the differential limiting torque also changes. As a result, when NF> NB turns, the front wheel side slips, and the NF <NB rear wheel side slips, the power distribution of the front and rear wheels is automatically controlled according to the difference in the action of the differential limiting torque. It is done.
[0060]
Therefore, the case of NF> NB will be described based on FIG. 6, FIG. 7, and FIG.
Under this condition, as shown in FIG. 8, when the input torque Ti is input in the counterclockwise direction of the first sun gear 39, the first and second pinions 41 and 42 rotate in the same direction, and the second sun gear is rotated. 40 and the carrier 38 also rotate in the same direction. Therefore, the friction torque Tf on the pinion side acts in the clockwise direction opposite to the pinion.
[0061]
Here, the torque and radius of each part are determined in the same manner as described above. Further, the tangential load P acting on the tooth surfaces of the first sun gear 39 of the first gear train and the first pinion 41, the separation load Fs1, the thrust load Ft1, the second sun gear 40 of the second gear train and the second gear. The tangential load P2, the separation load Fs2, and the thrust load Ft2 acting on the tooth surface of the second pinion 42 are assumed.
[0062]
Further, it has a friction coefficient μ1 between the side surfaces of the planetary pins 43, a sliding friction coefficient μ2, a friction torque Tf, a pinion inner radius re, and the friction coefficient μ2 at both end surfaces of the planetary pins 43. The outer radius rd of the surface that generates the friction, the number n of the contact surfaces, the module m1 of the first pinion 41, the torsion angle β1, the pressure angle α1, the module m2 of the second pinion 42, the torsion angle β2, and the pressure angle α2.
[0063]
Then
Fs1 = P · tanα1 / cosβ1
Ft1 = P · tanβ1
Is established, and the resultant force Np1 acting on the planetary pin 43 side is as follows.

Similarly,
Fs2 = P2 tan α2 / cos β2
Ft2 = P2 · tanβ2
Is established, and the resultant force Np2 acting on the planetarin pin 43 side is as follows.

The residual thrust force ΔFt generated in the first and second pinions 41 and 42 is as follows.

Accordingly, the friction torque Tf is the sum of the friction force generated by the two combined forces Np1 and Np2 and the friction force generated by the residual thrust force ΔFt as follows.

Next, the torque balance equation at the first and second pinions 41 and 42 is as follows.
Tf + P · rp1 = P2 · rp2 (15)
Further, when the friction torque Tf is added to the above equation (10), the result is as follows.

Here, as described above, each meshing pitch circle radius rs1, rs2, rp1, rp2 is replaced with the number of teeth Zs1, Zs2, Zp1, Zp2 of each gear, and the number of teeth is substituted for each number of teeth ( Zp1 = 24, Zp2 = 24, Zs1 = 30, Zs2 = 15) and the above equation (16) is
TB = 0.5Ti + 0.625Tf (17)
It becomes.
[0064]
Further, Ti = TF + TB. Substituting the above equation (16) for rearrangement results in the following.

Further, when the number of teeth Zs1, Zs2, Zp1, Zp2 is substituted and the number of teeth is substituted for each number of teeth, the above equation (18) becomes
TF = 0.5Ti−0.625Tf (19)
It becomes.
[0065]
Here, if μ1 = 0 and μ2 = 0, Tf = 0, and the values of the front and rear wheel side torques TF and TB indicate the same reference torque distribution as in the case of the above-described equal torque distribution function.
[0066]
Thus, it can be seen that under such conditions, a differential limiting torque Tf · rs 2 / rp 2 corresponding to the friction torque Tf is generated. The distribution of the front and rear wheel side torques TF and TB changes so that the rear wheel side is larger and the front wheel side is smaller by the differential limiting torque. Further, since the resultant forces Np1, Np2 and the residual thrust force ΔFt that generate the friction torque Tf are proportional to the input torque, they have an input torque proportional differential limiting function.
[0067]
On the other hand, the residual thrust force ΔFt is changed by the difference between the torsion angles β1 and β2 of the first and second pinions 41 and 42, and a needle bearing, bush or the like is used for the contact friction portion of the planetary pin 43. The friction coefficient μ1 can be changed. Thus, the value of the differential limiting torque can be set to various values together with the friction torque Tf.
[0068]
Subsequently, the case of NB> NF will be described. Under this condition, as shown in FIG. 9, the first and second pinions 41 and 42 revolve while rotating in the clockwise direction opposite to the first sun gear 39, and the friction torque Tf acts counterclockwise. . For this reason, the torque balance formula in the first and second pinions 41 and 42 is as follows.
Tf + P2 · rs2 = P · rp1 (20)
When calculated in the same manner as described above, the front and rear wheel side torques TF and TB are as follows.

Accordingly, even under this condition, the same differential limiting torque, Tf · rs 2 / rp 2, is generated. On the other hand, in this case, contrary to the above, torque is distributed so that the rear wheel side is small and the front wheel side is large by the differential limiting torque.
[0069]
Next, the rear wheel left / right driving force distribution unit 7 will be described in detail with reference to FIG.
The rear wheel left / right driving force distribution unit 7 is mainly composed of a differential limiting mechanism unit 44, a gear mechanism unit 45, and a clutch mechanism unit 46. The drive pinion 6 and the differential limiting mechanism 44 that transmit the driving force to the motor are housed in a differential carrier 47, and the clutch mechanism 46 is disposed on the side of the differential carrier 47 via the gear mechanism 45. The rear end portion of the differential carrier 47 is covered with a cover 48, and the gear mechanism portion 45 and the clutch mechanism portion 46 are covered with a cover 49.
[0070]
In the drive pinion 6, a shaft portion 6 a connected to the propeller shaft 5 is rotatably supported by a bearing in the differential carrier 47.
[0071]
First, the differential limiting mechanism 44 will be described.
The left drive shaft 14 is rotatably provided through a left side retainer 50 attached to the differential carrier 47. The right drive shaft 17 is coaxial with the left drive shaft 14, and the right drive shaft 17 is positioned on the right side of the differential carrier 47. It is provided so as to be rotatable through.
[0072]
A left differential case 51L is rotatably fitted to the outer periphery of the left drive shaft 14, and the left drive shaft 14 and the left differential case 51L are rotatably supported by the left side retainer 50 via a bearing. ing.
[0073]
In the left differential case 51L, one end portion (left end portion) of the right differential case 51R and a crown gear 52 meshed with the drive pinion 6 are both aligned and fixed at the center of rotation, and the right differential is fixed. The other end portion 51Ra of the case 51R is formed in a cylindrical shape, and is rotatably fitted to the outer periphery of a cylindrical portion 53a formed on the right side surface of the carrier 53, and the first gear shaft 54 in the gear mechanism portion 45. And is supported by the differential carrier 47 through a bearing.
[0074]
Further, the second gear shaft 55 in the gear mechanism portion 45 is spline-fitted inside the cylindrical portion 53 a of the carrier 53, and the right drive shaft 17 is located inside the second gear shaft 55. It is fitted freely.
[0075]
That is, the connecting portion between the right differential case 51R and the first gear shaft 54, the connecting portion between the carrier 53 and the second gear shaft 55, and the right drive shaft 17 are rotatable with respect to each other. The differential case 51, which is supported by the differential carrier 47 and includes the left differential case 51L and the right differential case 51R, to which the crown gear 52 is attached, is rotatably held in the differential carrier 47. Yes.
[0076]
In the differential case 51, the carrier 53 is rotatably arranged. In the carrier 53, the left drive shaft 14 and the right drive shaft 17 are inserted, and the carrier 53 is connected to the left drive shaft. 14 is splined to the tip.
[0077]
In the differential case 51, a first sun gear 56 having a large diameter is spline-coupled to a portion of the left differential case 51L through which the left drive shaft 14 is inserted, and a small diameter is provided to the tip portion of the right drive shaft 17. The second sun gear 57 is spline-coupled, and the first sun gear 56 is engaged with the first pinion 58 having a small diameter to form a first gear train, and the second sun gear 57 is a second gear having a large diameter. The second gear train is formed by meshing with the other pinion 59.
[0078]
The first pinion 58 and the second pinion 59 are formed integrally with a pinion member 60, and a plurality (for example, three) of the pinion members 60 are bearing on respective planetary pins 61 fixed to the carrier 53. It is pivotally supported via a shaft. Thrust load receiving washers are interposed between both ends of the pinion member 60 and the carrier 53.
[0079]
That is, the differential limiting mechanism 44 transmits the driving force from the drive pinion 6 to the first sun gear 56 via the crown gear 52 and the differential case 51, and the right drive from the second sun gear 57. A composite planetary gear type differential limiting device that outputs to the shaft 17 and outputs to the left drive shaft 14 from the carrier 53 is constituted.
[0080]
In the composite planetary gear type differential limiting device formed by the differential limiting mechanism 44, the first and second sun gears 56, 57 and a plurality of the first and first sun gears 56, 57 are arranged around the sun gears 56, 57. A differential function is provided by appropriately setting the number of teeth of the second pinions 58 and 59.
[0081]
Further, by properly setting the meshing pitch circle radius between the first and second sun gears 56 and 57 and the first and second pinions 58 and 59, the reference torque distribution is equal torque of 50:50 on the left and right. Has the function of distribution.
[0082]
Further, the first and second sun gears 56 and 57 and the first and second pinions 58 and 59 are, for example, helical gears, and the twist angles of the first gear train and the second gear train are set. The thrust load remains without canceling out the thrust load and the friction torque between the pinion end faces is separated by the meshing between the first and second pinions 58 and 59 and the surface of the planetary pin 61, and the tangential load. A differential limiting function is provided by setting so that the resultant force acts and frictional torque is generated to obtain a differential limiting torque proportional to the input torque.
[0083]
Note that a more detailed description of the composite planetary gear type differential limiting device formed by the differential limiting mechanism 44 is substantially the same as the description in the center differential device 3 and is therefore omitted.
[0084]
Next, the gear mechanism 45 will be described.
The first gear shaft 54 connected to the differential case 53 extends to the outside of the differential carrier 47 and is rotatably supported by the differential carrier 47 via a bearing, and is connected to the carrier 53. A first gear 62 is formed at the opposite end of the first gear 62.
[0085]
The second gear shaft 55 connected to the carrier 53 extends outside the differential carrier 47 so as to be rotatable in the first gear shaft 54 and has the first gear 62 at the end. The second gear 63 disposed adjacent to the outer side (right side) of the motor is splined.
[0086]
Further, a third gear 64 splined to the right drive shaft 17 is provided outside the second gear 63.
[0087]
The first, second, and third gears 62, 63, and 64 are arranged in parallel on the same rotational axis that is parallel to the rotational axis of the first, second, and third gears 62, 63, and 64, respectively. The fourth, fifth, and sixth gears 65, 66, and 67 are engaged with each other.
[0088]
That is, the gear mechanism section 45 includes a first gear train formed by the first gear 62 and the fourth gear 65, a second gear train formed by the second gear 63 and the fifth gear 66, and the like. The third gear train is composed of the third gear train 64 and the sixth gear 67.
[0089]
The gear ratios of the gear trains are determined by changing the number of teeth of the first, second, third, fourth, fifth and sixth gears 62, 63, 64, 65, 66, 67 to z1, z2, z3, As z4, z5, and z6, the first gear train is set to z4 / z1 = 0.9, the second gear train is set to z5 / z2 = 1, and the third gear train is set to z6 / z3 = 1. Only the gear ratio of the first gear train is set to the speed increasing ratio.
[0090]
The fourth gear 65 is integrally formed on one end side of the fourth gear shaft 68, and one end portion of the fourth gear shaft 68 is rotatably supported outside the differential carrier 47 via a bearing. And the other end is rotatably supported by the cover 49. A cylindrical clutch drum 69 having an opening on the fourth gear 65 side is fixed to the other end side of the fourth gear shaft 68.
[0091]
The fifth gear 66 is integrally formed on one end side of the fifth gear shaft 70 and serves as a second rotating member. The fifth gear shaft 70 is the fourth gear shaft 68. Is extended to the bottom 69a of the clutch drum 69 along the outer periphery of the clutch drum 69. A clutch hub 71 having a predetermined length is provided at the other end portion of the fifth gear shaft 70, and a plurality of plates 72 are alternately stacked between the clutch hub 71 and the clutch drum 69 so that a first hydraulic pressure is provided. A multi-plate clutch 73 is formed.
[0092]
Further, the sixth gear 67 is integrally formed on one end side of the sixth gear shaft 74 to form a third rotating member, and the sixth gear shaft 74 is the fifth gear shaft 70. Is extended to the plate end surface of the first hydraulic multi-plate clutch 73. A clutch hub 75 having a predetermined length is formed at the other end portion of the sixth gear shaft 74, and a plurality of plates 76 are alternately stacked between the clutch hub 75 and the clutch drum 69 to provide a second hydraulic pressure. A multi-plate clutch 77 is formed.
[0093]
The first hydraulic multi-plate clutch 73 can be pressed by a first piston 78 inserted from the clutch drum bottom 69a, and the first hydraulic chamber 79 for operating the first piston 78 is the above-mentioned. The hydraulic line 36 communicates with the first hydraulic line 36 a. The set hydraulic pressure applied to the first hydraulic chamber 79 in order to operate the first piston 78 is variable according to a value controlled by the left / right driving force distribution control unit 32.
[0094]
Similarly, the second hydraulic multi-plate clutch 77 can be pressed by a second piston 80 inserted from the clutch drum bottom 69a, and a second hydraulic chamber for operating the second piston 80. 81 communicates with the second hydraulic line 36b of the hydraulic line 36. The set hydraulic pressure applied to the second hydraulic chamber 81 in order to operate the second piston 80 is variable according to a value controlled by the left / right driving force distribution control unit 32.
[0095]
That is, the clutch mechanism portion 46 is formed by integrally forming the two clutches 73 and 77, and the gear ratio of the gear mechanism portion 45 is set as described above. When the hydraulic multi-plate clutch 73 is connected, a large amount of driving force is distributed to the left drive shaft 14 on the carrier 53 side, and when the second hydraulic multi-plate clutch 77 is connected, a large amount of driving force is distributed to the right drive shaft 17. It has come to be. Here, the hydraulic pressure value for connecting the hydraulic multi-plate clutches 73 and 77 is a value calculated by the left / right driving force distribution control unit 32, and the torque distribution amount is changed depending on the magnitude of the hydraulic pressure value.
[0096]
Further, driving force distribution performed by the gear mechanism 45 and the clutch mechanism 46 will be described. First, in order to improve the left-turning performance so that a large amount of driving force is distributed to the right drive shaft 17, the second hydraulic multi-plate clutch 77 on the right drive shaft 17 side is connected with a set hydraulic pressure. The driving force distributed to the carrier 53 side (left drive shaft 14 side) and the driving force distributed to the right drive shaft 17 side have a 50:50 relationship unless the differential limiting mechanism 44 operates. Therefore, if the driving force input to the rear-wheel left / right driving force distribution unit 7 is TB, each becomes TB / 2. Further, the driving force that is largely distributed from the second hydraulic multi-plate clutch 77 to the right drive shaft 17 is distributed via the third gear train. For this reason, with respect to the second hydraulic multi-plate clutch 77, the hydraulic pressure acting on the second hydraulic chamber 81, the dynamic friction coefficient of the friction surface (dynamic friction coefficient determined by the relative rotational speed of the friction surface), the number of friction surfaces (multiple If the clutch slip torque determined by the number of plate clutches × 2), the effective radius, etc. is Tk2, the driving force (right wheel side driving force) distributed to the right drive shaft 17 and the carrier 53 side are distributed. The driving force (left wheel side driving force) is as follows.
Right wheel side driving force = TB / 2 + (Tk2 / 2 × (z3 / z6)) × TB (25)
Left wheel side driving force = TB / 2− (Tk2 / 2 × (z3 / z6)) × TB (26)
Then, substituting the gear ratio z6 / z3 = 1 into the equations (25) and (26),
Right wheel side driving force = TB / 2 + Tk2 × (TB / 2) (25) ′
Left wheel side driving force = TB / 2−Tk2 × (TB / 2) (26) ′
It becomes.
[0097]
Further, in order to improve the right-turning performance so that a large amount of driving force is distributed to the carrier 53 side, the first hydraulic multi-plate clutch 73 on the carrier 53 side is connected with a set hydraulic pressure. The driving force that is largely distributed from the first hydraulic multi-plate clutch 73 to the carrier 53 is distributed through the second gear train. For this reason, regarding the first hydraulic multi-plate clutch 73, the hydraulic pressure that operates in the first hydraulic chamber 79, the dynamic friction coefficient of the friction surface (dynamic friction coefficient determined by the relative rotational speed of the friction surface), the number of friction surfaces (multiple If the clutch slip torque determined by the number of plate clutches x 2), effective radius, etc. is Tk1, the driving force (right wheel side driving force) distributed to the right drive shaft 17 and the carrier 53 side are distributed. The driving force (left wheel side driving force) is as follows.
Right wheel side driving force = TB / 2− (Tk1 / 2 × (z 2 / z 5)) × TB (27)
Left wheel side driving force = TB / 2 + (Tk1 / 2 × (z2 / z5)) × TB (28)
Then, substituting the gear ratio z5 / z2 = 1 into the above equations (27) and (28),
Right wheel side driving force = TB / 2 + Tk1 × (TB / 2) (27) ′
Left wheel side driving force = TB / 2−Tk1 × (TB / 2) (28) ′
It becomes.
[0098]
The two clutches of the clutch mechanism 46 may use an electromagnetic clutch or a variable transmission capacity coupling in addition to the hydraulic multi-plate clutch described above.
[0099]
On the other hand, the front wheel left / right driving force distribution unit 12 is configured such that a driving force is input from the front drive shaft 10 and the drive pinion 11 to the crown gear 52, and the structure thereof is the rear wheel left / right driving force distribution unit 7. The description is omitted because it is substantially the same.
[0100]
Next, the rear wheel side hydraulic control device 34 and the front wheel side hydraulic control device 35 will be described with reference to FIG. The rear wheel side hydraulic control device 34 has a pair of hydraulic paths, a hydraulic path for applying control hydraulic pressure to the first hydraulic chamber 79 through the hydraulic line 36a and a hydraulic path for applying to the second hydraulic chamber 81 through the hydraulic line 36b. The front wheel side hydraulic control device 35 is configured in substantially the same manner. Therefore, only the hydraulic path for applying the control hydraulic pressure to the first hydraulic chamber 79 through the hydraulic line 36a will be described below.
[0101]
The discharge pressure of the oil pump 83 driven by the motor 82 is regulated by the regulator valve 84 so as to generate a predetermined operating oil pressure and a lubricating oil pressure. An oil passage 85 for the operating oil pressure includes a clutch control valve 86, an oil pressure The first hydraulic multi-plate clutch 73 communicates with the first hydraulic chamber 79 side through a pipe line 36a.
[0102]
The oil passage 85 is connected to the control side of the duty solenoid valve 89 and the clutch control valve 86 by a pilot valve 87 and an oil passage 88.
[0103]
A duty signal from the left / right driving force distribution control unit 32 is output to the duty solenoid valve 89 to generate a duty pressure, and the clutch control valve 86 is operated with the duty pressure, whereby the first hydraulic pressure is obtained. The clutch hydraulic pressure of the multi-plate clutch 73 is controlled.
[0104]
Further, as shown in FIG. 11, the left / right driving force distribution control unit 32 mainly includes a road surface / running state determination unit 90, a hydraulic pressure calculation unit 91, and a hydraulic pressure setting unit 92, depending on the road surface / running state. Thus, an optimal driving force distribution amount between the front and rear wheels is calculated and a signal is output to the hydraulic control devices 35 and 34 on the front and rear wheels.
[0105]
The road surface / running state determination unit 90 receives the throttle opening sensor 26, the vehicle speed sensor 27, the rudder angle sensor 29, the longitudinal acceleration sensor 30, the lateral acceleration sensor 31, and the gear position signal, and the road surface based on these signals. The situation (whether or not on a low μ road, etc.) and the running state (high speed or low speed, sudden turning, high load or low load (acceleration state), presence / absence of slip state, etc.) are stored in advance. It is formed so as to be obtained by a map, a calculation formula or the like, and output to the hydraulic pressure calculation unit 91.
[0106]
Further, the hydraulic pressure calculation unit 91 selects and adds to the hydraulic multi-plate clutch to be operated based on the signal from the road surface / running state determination unit 90, based on a previously stored map, calculation formula, and the like. The hydraulic pressure value is calculated, and the result of the selection / calculation is output to the hydraulic pressure setting unit 92.
[0107]
The hydraulic pressure setting unit 92 is configured to output a signal from the hydraulic pressure calculation unit 91 to the corresponding hydraulic control device.
[0108]
Next, the operation of the above configuration will be described.
First, the driving force from the engine 1 is input from the automatic transmission 2 to the first sun gear 39 of the center differential device 3 via the transmission output shaft 2a.
[0109]
The first and second pinions 41 and 42 are distributed and transmitted to the second sun gear 40 and the carrier 38 that supports the first and second pinions 41 and 42, and the power of the second sun gear 40 is transmitted. Is transmitted to the rear wheel side via the rear drive shaft 4. The power of the carrier 38 is transmitted to the front wheels via the transfer drive gear 8, the transfer driven gear 9, and the front drive shaft 10, and travels by four-wheel drive.
[0110]
Therefore, for example, in a straight traveling of NF = NB where the front wheel side rotational speed is equal to the rear wheel side rotational speed, the second sun gear 40 and the carrier 38 rotate at the same speed in the same direction in the center differential device 3. The first and second pinions 41 and 42 stop rotating as a planet and rotate together.
[0111]
Thus, since the first and second pinions 41 and 42 and the carrier 38 are integrated, friction torque or the like is not generated between the two and the input torque Ti of the first sun gear 39 is reduced. On the other hand, the front wheel side torque TF of the carrier 38 and the rear wheel side torque TB of the second sun gear 40, if the gear specifications are set for equal torque distribution, the reference torque based on the gear specifications of the equal torque distribution function. If the distribution and the TF pair TB are set to only about 50 to 50 and the gear specifications are set for the unequal torque distribution, the TF vs TB is used for the reference torque distribution by the gear specifications of the unequal torque distribution function. Is set.
[0112]
Next, at the time of NF> NB turning or front wheel side slip where the front wheel side rotational speed is higher than the rear wheel side rotational speed, the first and second pinions 41 and 42 of the center differential device 3 are planetarily rotated to perform differential It works differentially with the gear specifications that have the function. For this reason, when turning, the difference in rotational speed between the front and rear wheels is absorbed and the vehicle turns smoothly.
[0113]
Along with the planetary rotation of the first and second pinions 41 and 42, a thrust load due to the difference in torsion angle acts on one end surface portion of the first and second pinions 41 and 42. Further, the combined force of the gear meshing point and the tangential load acts on the first and second pinions 41 and 42 and the planetary pin 43, thereby causing a friction torque opposite to the pinion rotation direction, Based on the differential limiting torque, it will occur.
[0114]
Under this condition, the differential limiting torque acts so as to impair the rotation of the carrier 38, so that the differential limiting torque moves to the rear wheel side, and the torque distribution becomes more deviated than the reference torque distribution. For this reason, turning ability and maneuverability during turning are improved, and slip is prevented at the time of front wheel slip when going straight.
[0115]
Further, when the rear wheel side rotational speed is larger than the front wheel side wheel rotational speed at the time of NB> NF rear wheel slip, the first and second pinions 41 and 42 of the center differential device 3 are similarly caused by the rotational speed difference between the front and rear wheels. Planetary rotation generates friction torque.
[0116]
By the way, under this condition, the differential limiting torque acts to promote the rotation of the carrier 38 and moves to the front wheel side, so that the torque distribution is larger on the front wheel side than the reference torque distribution, thereby preventing rear wheel slip. To do.
[0117]
Here, since the differential limiting torque by the planetary gear mechanism is generated in proportion to the input torque, it always has the same ratio with respect to the magnitude of the torque of the front and rear wheels, and the differential limiting function is always exerted at a constant ratio. Is done.
[0118]
As described above, the driving force distributed to one rear wheel side of the driving force distributed by the center differential device 3 is input to the rear wheel left / right driving force distribution unit 7 via the propeller shaft 5 and the drive pinion 6, and the crown It is input to the differential case 51 through the gear 52.
[0119]
First, the function of the differential limiting mechanism 44 will be described. The driving force input to the differential case 51 is input to the first sun gear 56, and the first and second pinions 58 and 59 to the second sun gear. 57 and the carrier 53 that supports the first and second pinions 58 and 59 are distributed and transmitted, and the power of the second sun gear 57 is transmitted to the right rear wheel 19 via the right drive shaft 17. The Further, the power of the carrier 53 is transmitted to the left rear wheel 16 via the left drive shaft 14 to drive.
[0120]
Therefore, for example, in a straight traveling of NL = NR where the left rear wheel rotational speed and the right rear wheel rotational speed are equal, the second sun gear 57 and the carrier 53 rotate at the same speed in the same direction, thereby The second pinions 58 and 59 stop rotating as a planet and rotate together.
[0121]
Thus, since the first and second pinions 58 and 59 and the carrier 53 are integrated, friction torque or the like is not generated between the two and the input torque of the first sun gear 56 is reduced. For the left rear wheel torque of the carrier 53 and the right rear wheel torque of the second sun gear 57, the reference torque distribution based on the gear specifications of the equal torque distribution function is set to approximately 50 to 50 only.
[0122]
Next, when the left rear wheel slip of NL> NR where the left rear wheel rotational speed is larger than the right rear wheel rotational speed, the first and second pinions 58 and 59 rotate in a planetary manner, and the gears having a differential function. Differential action by the origin. For this reason, when turning, the difference in rotational speed between the left and right rear wheels is absorbed and the vehicle turns smoothly.
[0123]
Along with the planetary rotation of the first and second pinions 58 and 59, a thrust load due to the difference in torsion angle acts on one end face portion of the first and second pinions 58 and 59. Further, the combined force of the gear meshing point and the tangential load acts on the first and second pinions 58 and 59 and the planetary pin 61, thereby causing a friction torque opposite to the pinion rotation direction, Based on the differential limiting torque, it will occur.
[0124]
Under this condition, the differential limiting torque acts so as to impair the rotation of the carrier 53, so that the differential limiting torque moves to the right rear wheel side, and the torque distribution is more deviated from the reference torque distribution to the right rear wheel. Become. This prevents slipping when the left rear wheel slips straight ahead.
[0125]
Further, when the right rear wheel slip is larger than the left rear wheel rotation speed and the right rear wheel slips with NR> NL, the first and second pinions 58 and 59 are similarly planetarily rotated due to the difference in the rotation speed between the left and right rear wheels. Generate friction torque.
[0126]
By the way, under this condition, the differential limiting torque acts to encourage the rotation of the carrier 53 and moves to the left rear wheel side, so that the torque distribution is greater on the left rear wheel than the reference torque distribution, resulting in a right rear wheel. Prevent wheel slip.
[0127]
Here, since the differential limiting torque by the planetary gear mechanism is generated in proportion to the input torque, it always has the same ratio with respect to the magnitude of the torque of the left and right rear wheels, and the differential limiting function is always at a constant ratio. Demonstrated.
[0128]
Next, the gear mechanism portion 45 and the clutch mechanism portion 46 will be described. When the differential case 51 is rotated, the first gear shaft 54 fixed to the differential case 51 is rotated and the first gear 62 is moved. The fourth gear 65 that is rotated and meshed with the first gear 61 is rotated, and the fourth gear shaft 68 and the clutch drum 69 are rotated.
[0129]
Further, by the rotation of the carrier 53 on the left wheel output side, the second gear shaft 55 fixed to the carrier 53 is rotated and the second gear 63 is rotated and meshed with the second gear 63. The fifth gear 66 is rotated, and the fifth gear shaft 70 and the clutch hub 71 are rotated.
[0130]
Further, the third gear 64 is rotated together with the rotation of the right drive shaft 17, the sixth gear 67 meshed with the third gear 64 is rotated, and the sixth gear shaft 74 and the clutch hub 75 are rotated. Is done.
[0131]
On the other hand, in the rear wheel side hydraulic control device 34, the oil pump 83 is driven by the motor 82, and the operating hydraulic pressure by the regulator valve 84 is guided to the duty solenoid valve 89 and the clutch control valve 86.
[0132]
In the left / right driving force distribution control unit 32, a throttle opening sensor 26, a vehicle speed sensor 27, a rudder angle sensor 29, a longitudinal acceleration sensor 30, a lateral acceleration sensor 31, and a gear position signal are input and processed, and front and rear left and right wheels are processed. The optimal driving force distribution amount is calculated.
[0133]
When the vehicle distributes a large amount of driving force to the left side on the rear wheel side in a right turn state, the first hydraulic chamber 79 is transmitted from the left / right driving force distribution control unit 32 to the rear wheel side hydraulic control device 34. A signal is sent to apply hydraulic pressure at the set pressure calculated on the side.
[0134]
As a result, the hydraulic pressure is applied to the first hydraulic chamber 79 via the hydraulic line 36a, the first piston 78 is operated, the first hydraulic multi-plate clutch 73 is connected with the set pressure, and the differential case The driving force from 51 is distributed to the carrier 53 and the left drive shaft 14 via the first hydraulic multi-plate clutch 73 and the second gear 63.
[0135]
On the other hand, when the vehicle distributes a large amount of driving force to the right side on the rear wheel side in a left turn state, the second hydraulic chamber 81 side from the left / right driving force distribution control unit 32 to the rear wheel side hydraulic control device 34 A signal is sent to apply the hydraulic pressure at the set pressure calculated in step (b).
[0136]
As a result, the hydraulic pressure is applied to the second hydraulic chamber 81 via the hydraulic line 36b, the second piston 80 is operated, the second hydraulic multi-plate clutch 77 is connected at the set pressure, and the differential is
The driving force from the case 51 is distributed to the right drive shaft 17 via the second hydraulic multi-plate clutch 77 and the third gear 64.
[0137]
Further, the driving force distributed to the other front wheel side of the driving force distributed by the center differential device 3 is input to the front wheel left / right driving force distribution unit 12 via the front drive shaft 10 and the drive pinion 11, and the crown gear 52. To the differential case 51.
[0138]
The front wheel left / right driving force distribution unit 12 also has a front wheel side differential function similar to the rear wheel left / right driving force distribution unit 7, and the front wheel receives a signal from the left / right driving force distribution control unit 32. When a large amount of driving force is distributed to the left side of the front wheel side when the vehicle is turning right, the hydraulic pressure is supplied to the first hydraulic chamber 79 via the hydraulic line 37a. In addition, when the vehicle distributes a large amount of driving force to the right side on the front wheel side in a left turn state, hydraulic pressure is applied to the second hydraulic chamber 81 via the hydraulic line 37b.
[0139]
As described above, according to the first embodiment of the present invention, the right and left driving force distribution device is prevented from increasing the width dimension in the left and right wheel direction, and the intersection angle of the universal joint disposed between the left and right wheels and the axle shaft. Can be reduced, and durability and reliability can be improved.
[0140]
Further, the left and right driving force distribution device has a small number of components, can use many conventional devices, is compatible with conventional devices, is compact, and can be manufactured at low cost.
[0141]
Further, the left and right driving force distribution device is excellent in in-vehicle performance because it has a compact width dimension in the left and right wheel directions, and can ensure a clearance during maintenance and interference with components of the suspension and exhaust system members.
[0142]
Moreover, since the left and right driving force distribution device is configured to bypass the driving force from the differential case to both output sides, the whole is compact.
[0143]
Furthermore, the left / right driving force distribution device employs a driving force distribution mechanism consisting of three rows of gear trains, so the driving force distribution can be adjusted by setting the gear ratio and can easily be adapted to the personality and purpose of the vehicle. It is.
[0144]
Further, the two sets of hydraulic multi-plate clutches of the left and right driving force distribution device have an integrated structure, and are lightweight and compact.
[0145]
In addition, since the differential limiting mechanism has a torque sensitive differential limiting function with the frictional force acting on the rotational friction surfaces such as gear specifications (torsion angle, pressure angle, etc.) and pinions and carriers, planetary pins, etc. An apparatus having both the differential limiting function and active left / right driving force distribution is realized.
[0146]
In addition, the center differential device has a simple structure, a small number of parts, is lightweight and compact, and thus has excellent workability and assemblability, and is advantageous in terms of vibration noise in the power transmission system.
[0147]
Furthermore, both the center differential device and the left / right driving force distribution device are lightweight and compact, and can be easily integrated, and a lightweight and compact integrated unit can be realized.
[0148]
Also, the center differential device can set the number of teeth so that the reference torque distribution is distributed to the front and rear wheels at a ratio of 50:50, and the input torque proportional differential limiting torque is applied to the front wheels or the rear wheels. It can move according to the road surface conditions, prevent slipping of the vehicle, prevent behavior such as ensuring driving force and swinging the vehicle, and improve running performance. In addition, it is easy to control the attitude of the vehicle with respect to the accelerator operation, and it is possible to enjoy sporty driving with a good response.
[0149]
Next, FIG. 12 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 2 of the present invention. In the second embodiment of the present invention, the position of the clutch mechanism portion of the rear wheel left / right driving force distribution portion of the first embodiment of the present invention is changed.
[0150]
That is, as shown in the figure, the first hydraulic multi-plate clutch 73 related to the second gear 63 is formed on the left side (the differential case 51 side) of the fourth gear 65, and further the third hydraulic clutch 73 is formed on the left side. A second hydraulic multi-plate clutch 77 relating to the gear 64 is formed to constitute a two-piece clutch mechanism.
[0151]
As described above, by changing the position of the clutch mechanism portion, it is possible to further shorten the left-right direction length of the left-right driving force distribution portion.
[0152]
Next, FIG. 13 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 3 of the present invention. In the third embodiment of the present invention, the gear ratio of the first gear train of the gear mechanism section of the rear wheel left / right driving force distribution section of the first embodiment of the present invention is changed.
[0153]
In the rear wheel left / right driving force distribution unit shown in the figure, the gear ratio of each gear train is set as follows. The first gear train is z4 / z1 = 1 / 0.9, the second gear train is z5 / z2 = 1, the third gear train is z6 / z3 = 1, and the first gear train Only the gear ratio is set as the reduction ratio. For this reason, the operation | movement by connection of each clutch is performed contrary to the operation | movement of Embodiment 1 of the said invention.
[0154]
That is, when a large amount of driving force is distributed to the left side on the rear wheel side in a right turn state, the second hydraulic chamber 81 is transferred from the left / right driving force distribution control unit 32 to the rear wheel side hydraulic control device 34. A signal is sent to apply hydraulic pressure at the set pressure calculated on the side.
[0155]
As a result, the hydraulic pressure is applied to the second hydraulic chamber 81 via the hydraulic line 36b, the second piston 80 is operated, the second hydraulic multi-plate clutch 77 is connected at the set pressure, and the differential case The driving force from 51 decreases to the right drive shaft 17 side and is distributed more to the carrier 53 side.
[0156]
On the other hand, when the vehicle distributes a large amount of driving force to the right side on the rear wheel side in the left turn state, the first hydraulic chamber 79 side from the left / right driving force distribution control unit 32 to the rear wheel side hydraulic control device 34 A signal is sent to apply the hydraulic pressure at the set pressure calculated in step (b).
[0157]
As a result, the hydraulic pressure is applied to the first hydraulic chamber 79 via the hydraulic line 36a, the first piston 78 is operated, the first hydraulic multi-plate clutch 73 is connected with the set pressure, and the differential case The driving force from 51 decreases toward the carrier 53 and is distributed more to the right drive shaft 17 side.
[0158]
Next, FIG. 14 is an enlarged skeleton diagram of a rear-wheel left / right driving force distribution unit according to Embodiment 4 of the present invention. In the fourth embodiment of the present invention, the position of the clutch mechanism portion of the rear-wheel left / right driving force distribution portion of the third embodiment of the present invention is changed.
[0159]
That is, as shown in the figure, the first hydraulic multi-plate clutch 73 related to the second gear 63 is formed on the left side (the differential case 51 side) of the fourth gear 65, and further the third hydraulic clutch 73 is formed on the left side. A second hydraulic multi-plate clutch 77 relating to the gear 64 is formed to constitute a two-piece clutch mechanism.
[0160]
As described above, by changing the position of the clutch mechanism portion, it is possible to further shorten the left-right direction length of the left-right driving force distribution portion.
[0161]
Next, FIG. 15 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 5 of the present invention. In the fifth embodiment of the present invention, the gear mechanism portion of the second embodiment of the present invention is formed separately.
[0162]
That is, as shown in the figure, the second gear 63 is not fixed to the carrier 53 but is fixed to the left drive shaft 14. Therefore, the first gear train and the third gear train are separated from the differential case 51 on the right side, and the second gear train is formed on the left side.
[0163]
Next, FIG. 16 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 6 of the present invention. In the sixth embodiment of the present invention, the gear mechanism portion of the fourth embodiment of the present invention is formed separately.
[0164]
That is, as shown in the figure, the second gear 63 is not fixed to the carrier 53 but is fixed to the left drive shaft 14. Therefore, the first gear train and the third gear train are separated from the differential case 51 on the right side, and the second gear train is formed on the left side.
[0165]
Next, FIG. 17 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 7 of the present invention. In the seventh embodiment of the present invention, driving force distribution is performed by two gear trains provided on one output side.
[0166]
That is, as shown in the drawing, the first gear 62 is fixed to the differential case 51, and the second gear 95 and the third gear 96 are fixed to the right drive shaft 17. The second and third gears are meshed with fourth, fifth, and sixth gears 65, 97, and 98 on the same rotational axis.
[0167]
The second gear 95 is a gear having a larger diameter than the first gear 62, and the third gear 96 is a gear having a smaller diameter than the first gear 62. A first gear train composed of the first gear 62 and the fourth gear 65; a second gear train composed of the second gear 95 and the fifth gear 97; The third gear train composed of the third gear 96 and the sixth gear 98 has a gear ratio of the first, second, third, fourth, fifth and sixth gears 62, 95, 96, 65, 97, The number of teeth of 98 is z1, z2, z3, z4, z5, z6, respectively, the first gear train is z4 / z1 = 0.9, and the second gear train is z5 / z2 = 0.9 × 0. .9, the third gear train is set to z 6 / z 3 = 1, and when the gear ratios are arranged in descending order, 1 (third gear train), 0.9 (first gear train), 0. At 9 × 0.9 (second gear train), the step ratio is constant at 0.9. This value may be set to another value.
[0168]
The first gear train and the second gear train are connected by a first hydraulic multi-plate clutch 99 so that power can be transmitted, and the first gear train and the third gear train are configured. The space is connected by the second hydraulic multi-plate clutch 100 so that power can be transmitted.
[0169]
The capacity of the two sets of hydraulic multi-plate clutches may be set to unequal values in advance according to the gear ratio, and one of the multi-plate clutches can be reduced in size.
[0170]
With this configuration, when the vehicle distributes a large amount of driving force to the left side on the rear wheel side in a right turn state, the left and right driving force distribution control unit 32 controls the rear wheel side hydraulic control device 34 to A signal is sent to apply the hydraulic pressure at the set pressure calculated to the first hydraulic chamber 79 side.
[0171]
As a result, the hydraulic pressure is applied to the first hydraulic chamber 79 via the hydraulic line 36a, the first piston 78 is operated, the first hydraulic multi-plate clutch 73 is connected with the set pressure, and the differential case The driving force from 51 decreases to the right drive shaft 17 side and is distributed more to the carrier 53 side.
[0172]
On the other hand, when the vehicle distributes a large amount of driving force to the right side on the rear wheel side in a left turn state, the second hydraulic chamber 81 side from the left / right driving force distribution control unit 32 to the rear wheel side hydraulic control device 34 A signal is sent to apply the hydraulic pressure at the set pressure calculated in step (b).
[0173]
As a result, the hydraulic pressure is applied to the second hydraulic chamber 81 via the hydraulic line 36b, the second piston 80 is operated, the second hydraulic multi-plate clutch 77 is connected at the set pressure, and the differential case A large amount of driving force from 51 is distributed to the right drive shaft 17 side.
[0174]
Also according to the seventh embodiment of the present invention, the left / right driving force distribution device sets the gear ratios of the three gear trains to a predetermined relationship, so that the drive plates and the driven plates of the two hydraulic multi-plate clutches are relatively Since the rotation difference can be made the same, a range in which the same friction characteristic (relationship between speed and dynamic friction coefficient) can be obtained can be used, and the control accuracy can be increased.
[0175]
As shown in the embodiments of the inventions described above, the positions of the gear trains constituting the gear mechanism and the positions of the clutch mechanisms are variously set and are positions other than those shown in the embodiments of the inventions. May be. In addition, the gear ratio of each gear train of the gear mechanism described in the embodiments of each invention may be set to other values.
[0176]
Vehicles other than those described in the above embodiments of the invention, vehicles that perform left and right driving force distribution among front and rear wheels of FF (front engine / front drive) vehicles, RR (rear engine / rear drive) vehicles, and 4WD vehicles The present invention can also be applied to the above.
[0177]
Further, the center differential device when applied to a 4WD vehicle may be other than the above-described one.
[0178]
【The invention's effect】
As described above, according to the present invention, it is possible to prevent an increase in the length of the left and right output shafts, and to reduce the intersection angle of the universal joint disposed between the left and right wheels and the axle shaft. Excellent in-vehicle performance, such as interference with exhaust system members and ensuring clearance during maintenance, small number and small number of components, mounting compatibility with conventional differentials, and advantageous in manufacturing cost. Also, it has a differential limiting function, and it is easy to adjust and set the bypass torque, has high control accuracy, and is excellent in durability and reliability.
[Brief description of the drawings]
FIG. 1 is an explanatory diagram showing an overall schematic configuration of a 4WD vehicle according to a first embodiment of the present invention.
FIG. 2 is an enlarged cross-sectional view of a rear wheel left / right driving force distribution unit according to Embodiment 1 of the present invention;
FIG. 3 is a schematic diagram of each part for explaining a differential function of the center differential device according to the first embodiment of the present invention;
FIG. 4 is an operation explanatory diagram when the first sun gear is fixed according to the first embodiment of the present invention.
FIG. 5 is an operation explanatory diagram when the second sun gear is fixed according to the first embodiment of the present invention.
FIG. 6 is a schematic diagram of each part for explaining a power distribution function and a differential limiting function of the center differential device according to the first embodiment of the present invention;
FIG. 7 is an explanatory diagram of a load generated by each gear according to the first embodiment of the present invention.
FIG. 8 illustrates a case where the front wheel side rotational speed is larger than the rear wheel side rotational speed according to the first embodiment of the present invention.
FIG. 9 is an explanatory diagram when the front wheel side rotational speed is smaller than the rear wheel side rotational speed according to the first embodiment of the present invention;
FIG. 10 is an explanatory diagram of the configuration of the hydraulic control device for left / right driving force distribution according to Embodiment 1 of the present invention;
FIG. 11 is a functional block explanatory diagram of a left / right driving force distribution control unit according to the first embodiment of the present invention;
FIG. 12 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 2 of the present invention;
FIG. 13 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 3 of the present invention;
FIG. 14 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 4 of the present invention;
FIG. 15 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 5 of the present invention;
FIG. 16 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 6 of the present invention;
FIG. 17 is an enlarged skeleton diagram of a rear wheel left / right driving force distribution unit according to Embodiment 7 of the present invention;
[Explanation of symbols]
1 engine
2 Automatic transmission
3 Center differential unit
4 Rear drive shaft
5 Propeller shaft
6 Drive pinion
7 Rear wheel left / right driving force distribution part
8 Transfer drive gear
9 Transfer driven gear
10 Front drive shaft
11 Drive pinion
12 Front wheel left / right driving force distribution part
14 Left drive shaft
16 Left rear wheel
17 Right drive shaft
19 Right rear wheel
20 Left drive shaft
21 Front left wheel
22 Right drive shaft
23 Right front wheel
26 Throttle opening sensor
27 Vehicle speed sensor
29 Rudder angle sensor
30 Longitudinal acceleration sensor
31 Lateral acceleration sensor
32 Left and right driving force distribution control unit
34 Rear wheel hydraulic control device
35 Front wheel side hydraulic control device
44 Differential limiting mechanism
45 Gear mechanism
46 Clutch mechanism
51 differential case
52 Crown Gear
53 Carrier
54 First gear shaft
55 Second gear shaft
56 First sun gear
57 Second sun gear
58 First Pinion
59 Second Pinion
60 Pinion member
61 Planetary pins
62 First gear
63 Second gear
64 third gear
65 fourth gear
66 fifth gear
67 6th gear
68 Fourth gear shaft
69 Clutch drum
70 fifth gear shaft
71 Clutch hub
72 plates
73 First hydraulic multi-plate clutch
74 6th gear shaft
75 Clutch hub
76 plates
77 Second hydraulic multi-plate clutch
78 First piston
79 First hydraulic chamber
80 second piston
81 Second hydraulic chamber

Claims (8)

The first sun gear on the driving force input side is meshed with the first pinion to form a first gear train, and the second sun gear on the output side of either the left wheel side or the right wheel side is connected to the first sun gear. The second pinion integral with the first pinion to form a second gear train, and the first and second pinions are pivotally supported by the carrier on the other output side, A differential limit proportional to the input torque between the left and right wheels by a frictional force obtained by applying a separation load acting on the gear meshing point of the second gear train to the shaft support portions of the first and second pinions. A first gear for generating a reference rotational speed for the first rotating member under the condition that there is no differential between the left and right wheels on the input side of the driving force; What is the reference rotational speed on the condition that there is no difference between the left and right wheels on either the output side or the other output side? A second gear that rotates the second rotating member at a second rotational speed is provided, and the reference rotation is performed on the condition that there is no differential between the left and right wheels on either the one output side or the other output side A third gear for rotating the third rotating member at a third rotational speed different from the speed is provided, and the first rotating member, the second rotating member, and the third rotating member are output to the both outputs. A clutch mechanism in which the first rotating member and the second rotating member are variable in transmission capacity, the first rotating member, and the third rotating shaft. A left-right driving force distribution device for a vehicle, wherein two sets of clutch mechanisms, each of which has a variable transmission capacity, are integrally formed on the rotational axis of each of the rotating members. A frictional force obtained by applying a difference in thrust load acting on a gear meshing point of the first gear train and the second gear train to one end face of the first and second pinions; and the first gear train. The frictional force obtained by applying the combined force of the separation load and tangential load acting on the gear meshing point of the second gear train and the second gear train to the shaft support portions of the first and second pinions, The vehicle left-right driving force distribution device according to claim 1, further comprising a differential limiting device that generates a differential limiting torque proportional to the input torque between the wheels. The left and right driving force distribution device for a vehicle according to claim 1 or 2, wherein the two sets of clutch mechanisms are formed of at least one of a hydraulic multi-plate clutch, an electromagnetic clutch, and a transmission capacity variable type coupling. 4. The left / right driving force distribution for a vehicle according to claim 1, wherein the transmission capacity by the two sets of clutch mechanisms is variably set according to the running state of the vehicle and the road surface condition. apparatus. The second gear of the second rotating member under the condition that the second gear is provided on the one output side, the third gear is provided on the other output side, and there is no differential between the left and right wheels. The left-right drive for vehicles according to any one of claims 1, 2, 3, and 4, wherein the rotation speed of the third rotation member and the third rotation speed of the third rotation member are set to the same value. Power distribution device. Both of the second gear and the third gear are provided on either the one output side or the other output side, and the second rotating member is operated under the condition that there is no differential between the left and right wheels. The second rotation speed and the third rotation speed of the third rotation member are set such that the rotation speed on one side is larger than the reference rotation speed and the rotation speed on the other side is smaller than the reference rotation speed. The left and right driving force distribution device for a vehicle according to any one of claims 1, 2, 3, and 4, characterized in that: The ratio of the rotation speed of the third gear and the third rotation speed of the third rotation member under the condition that there is no differential between the left and right wheels, the rotation speed of the first gear, and the first rotation The step ratio between the ratio of the reference rotational speed of the member and the ratio of the rotational speed of the second gear and the second rotational speed of the second rotating member is set constant. 7. The vehicle left-right driving force distribution device according to claim 6. 8. The left and right driving force distribution device for a vehicle according to claim 7, wherein the two sets of clutch mechanisms are formed with a predetermined transmission capacity difference.

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