JP3608442B2 - Vehicle damping coefficient control device - Google Patents

Description

【００２８】
車体のロール運動を減衰させるパラメータとしてＣｎ ＝ＷＣ／２とすると、上記式２は下記の式３の如く表わされる。
【００２９】
【数２】

【００３０】
また図６に示された実際の車輌の二輪モデルに於ける上下方向の力の釣り合い及び重心１１８の周りの力の釣り合いよりそれぞれ下記の式４及び式５が成立する。
【００３１】
【数３】

【００３２】
上記式１及び式４より下記の式６が成立する。
【００３３】
【数４】

【００３４】
またここでＣｍ ＝Ｃｎ ／Ｌとすると、上記式３及び式５より下記の式７が成立する。
【００３５】
【数５】

【００３６】
ここで図７に示された仮想モデルに於いてショックアブソーバ１２２により発生される上下力を下記の式８に従ってＴと置くと、上記式６、７及び下記の式８より下記の式９〜１１が成立する。
【００３７】
【数６】

【００３８】
【数７】

【００３９】
式９＋式１１より旋回内輪のショックアブソーバの減衰係数Ｃｉｎを以下の如く求めることができる。
【００４０】
【数８】

【００４１】
また上記式１２を式９に代入して旋回外輪のショックアブソーバの減衰係数Ｃｏｕｔ を以下の如く求めることができる。
【００４２】
【数９】

【００４３】
更に上記式１２及び式１３を整理して旋回内輪及び旋回外輪のショックアブソーバの減衰係数Ｃｉｎ及びＣｏｕｔ はそれぞれ下記の式１４及び式１５の如く表わされる。
【００４４】
【数１０】

【００４５】
尚旋回内輪側及び旋回外輪側のショックアブソーバにより発生される減衰力はそれぞれ下記の式１６及び式１７の如く求められる。
【００４６】
【数１１】

【００４７】
また同様の考え方に基づき、車体ロール量の減少過程に於いては、車輌の旋回外側に仮想のショックアブソーバ１２２及び１２４が配設された仮想モデルに基づき、旋回内輪側及び旋回外輪側のショックアブソーバの減衰係数Ｃｉｎ及びＣｏｕｔ をそれぞれ下記の式１８及び式１９の如く制御することにより、車輌の重心１１８の高さを低下させ、車輌の旋回終期に於ける運動性能を向上させることができる。
【００４８】
【数１２】

【００４９】
従って本発明の一つの好ましい態様によれば、上記請求項１の構成に於いて、車体ロール量の増大過程に於いては旋回内側のショックアブソーバの減衰係数Ｃｉｎ及び旋回外側のショックアブソーバの減衰係数Ｃｏｕｔ はそれぞれ上記式１４及び式１５に従って演算されるよう構成される（好ましい態様１）。
【００５０】
また本発明の他の一つの好ましい態様によれば、上記請求項１の構成に於いて、車体ロール量の減少過程に於いては旋回外側のショックアブソーバの減衰係数が旋回内側のショックアブソーバの減衰係数よりも高く制御されるよう構成される（好ましい態様２）。
【００５１】
また本発明の他の一つの好ましい態様によれば、上記好ましい態様２の構成に於いて、車体ロール量の減少過程に於いては旋回内側のショックアブソーバの減衰係数Ｃｉｎ及び旋回外側のショックアブソーバの減衰係数Ｃｏｕｔ はそれぞれ上記の式１８及び式１９に従って演算されるよう構成される（好ましい態様３）。
【００５２】
また本発明の他の一つの好ましい態様によれば、上記請求項２の構成に従い、上記好ましい態様１の構成に於いて、車輌モデルは前輪側の車輌モデルと後輪側の車輌モデルとよりなるよう構成される（好ましい態様４）。
【００５３】
また図８に示されている如く、前輪側及び後輪側の車輌モデルについてのＬ、Ｗ、Ｔ、Ｃｇ 、ＣをそれぞれＬｆ 及びＬｒ 、Ｗｆ 及びＷｒ 、Ｔｆ 及びＴｒ 、Ｃｇｆ及びＣｇｒ、Ｃｆ 及びＣｒ とし、旋回内側前輪及び旋回外側前輪のストローク速度をそれぞれＸｆｉｎｄ及びＸｆｏｕｔｄ とし、旋回内側後輪及び旋回外側後輪のストローク速度をそれぞれＸｒｉｎｄ及びＸｒｏｕｔｄ とし、Ｔｆ 及びＴｒ をそれぞれ下記の式２０及び式２１により表される値として、車体ロール量の増大過程に於いては旋回内側前輪のショックアブソーバの減衰係数Ｃｆｉｎ 及び旋回外側前輪のショックアブソーバの減衰係数Ｃｆｏｕｔはそれぞれ下記の式２２及び式２３に従って演算され、旋回内側後輪のショックアブソーバの減衰係数Ｃｒｉｎ 及び旋回外側後輪のショックアブソーバの減衰係数Ｃｒｏｕｔはそれぞれ下記の式２４及び式２５に従って演算されることが好ましい。
【００５４】
【数１３】

【００５５】
【数１４】

【００５６】
従って本発明の他の一つの好ましい態様によれば、上記好ましい態様４の構成に於いて、車体ロール量の増大過程に於いては旋回内側前輪のショックアブソーバの減衰係数Ｃｆｉｎ 及び旋回外側前輪のショックアブソーバの減衰係数Ｃｆｏｕｔはそれぞれ上記式２２及び式２３に従って演算され、旋回内側後輪のショックアブソーバの減衰係数Ｃｒｉｎ 及び旋回外側後輪のショックアブソーバの減衰係数Ｃｒｏｕｔはそれぞれ上記式２４及び式２５に従って演算されるよう構成される（好ましい態様５）。
【００５７】
同様に本発明の他の一つの好ましい態様によれば、上記好ましい態様３の構成に於いて、車輌モデルは前輪側の車輌モデルと後輪側の車輌モデルとよりなるよう構成される（好ましい態様６）。
【００５８】
また本発明の他の一つの好ましい態様によれば、上記好ましい態様６の構成に於いて、車体ロール量の減少過程に於いては旋回内側前輪のショックアブソーバの減衰係数Ｃｆｉｎ 及び旋回外側前輪のショックアブソーバの減衰係数Ｃｆｏｕｔはそれぞれ下記の式２６及び式２７に従って演算され、旋回内側後輪のショックアブソーバの減衰係数Ｃｒｉｎ 及び旋回外側後輪のショックアブソーバの減衰係数Ｃｒｏｕｔはそれぞれ下記の式２８及び式２９に従って演算されるよう構成される（好ましい態様７）。
【００５９】
【数１５】

【００６０】
また本発明の他の一つの好ましい態様によれば、上記請求項３の構成に於いて、車輌の状態量を検出する手段は車輌の基準ヨーレートと車輌の実ヨーレートとの偏差を車輌の旋回挙動として検出するよう構成される（好ましい態様８）。
【００６１】
また本発明の他の一つの好ましい態様によれば、上記請求項３の構成に於いて、車輌の状態量を検出する手段は車輌のスリップ角及びその変化率の関係を車輌の旋回挙動として検出するよう構成される（好ましい態様９）。
【００６２】
また本発明の他の一つの好ましい態様によれば、上記請求項４の構成に於いて、車輌の状態量を検出する手段は車速、操舵角、操舵角速度に基づき車輌の横加加速度を推定するよう構成される（好ましい態様１０）。
【００６３】
また本発明の他の一つの好ましい態様によれば、上記請求項５の構成に於いて、上記請求項２の構成に従い、距離設定手段は左右前輪の差圧の大きい方の値に応じて前輪の車輌モデルの所定の距離を可変設定し、左右後輪の差圧の大きい方の値に応じて後輪の車輌モデルの所定の距離を可変設定するよう構成される（好ましい態様１１）。
【００６４】
また本発明の他の一つの好ましい態様によれば、上記請求項６の構成に於いて、車輌の旋回挙動がオーバステア側へのステア特性変化の挙動であるときには前輪側のロール剛性を増大させ若しくは後輪側のロール剛性を低下させるよう前輪の車輌モデルと後輪の車輌モデルとの間の所定の距離の比を可変設定するよう構成される（好ましい態様１２）。
【００６５】
また本発明の他の一つの好ましい態様によれば、上記請求項６の構成に於いて、車輌の旋回挙動がアンダステア側へのステア特性変化の挙動であるときには前輪側のロール剛性を低下させ若しくは後輪側のロール剛性を増大させるよう前輪の車輌モデルと後輪の車輌モデルとの間の所定の距離の比を可変設定するよう構成される（好ましい態様１３）。
【００６６】
また本発明の他の一つの好ましい態様によれば、上記請求項６の構成に於いて、車輌の状態量を検出する手段はばね上のピッチ運動状態量を検出するよう構成される（好ましい態様１４）。
【００６７】
また本発明の他の一つの好ましい態様によれば、上記好ましい態様１４の構成に於いて、ばね上のピッチ運動状態量は車輌の加減速度であるよう構成される（好ましい態様１５）。
【００６８】
また本発明の他の一つの好ましい態様によれば、上記好ましい態様１５の構成に於いて、車輌の加減速度は運転者による加減速操作量に基づき推定される推定加減速度であるよう構成される（好ましい態様１６）。
【００６９】
また本発明の他の一つの好ましい態様によれば、上記好ましい態様１６の構成に於いて、運転者による加減速操作量はブレーキペダルのストロークであるよう構成される（好ましい態様１７）。
【００７０】
また本発明の他の一つの好ましい態様によれば、上記好ましい態様１６の構成に於いて、運転者による加減速操作量はスロットル開度速度であるよう構成される（好ましい態様１８）。
【００７１】
【発明の実施の形態】
以下に添付の図を参照しつつ、本発明を幾つかの好ましい実施形態について詳細に説明する。
【００７２】
第一の実施形態
図１は本発明による減衰係数制御装置の第一の好ましい実施形態を示す概略構成図である。
【００７３】
図１に於て、１０ＦＬ及び１０ＦＲはそれぞれ車輌１２の左右の前輪を示し、１０ＲＬ及び１０ＲＲはそれぞれ左右の後輪を示している。操舵輪である左右の前輪１０ＦＬ及び１０ＦＲは運転者によるステアリングホイール１４の転舵に応答して駆動されるラック・アンド・ピニオン式のパワーステアリング装置１６によりタイロッド１８Ｌ及び１８Ｒを介して操舵される。
【００７４】
ばね下としての各車輪１０ＦＬ〜１０ＲＲとばね上としての車体２０との間にはそれぞれ減衰係数可変式のショックアブソーバ２２ＦＬ〜２２ＲＲが配設されており、各ショックアブソーバの減衰係数Ｃｉ（ｉ＝ｆｌ、ｆｒ、ｒｌ、ｒｒ）は後述の如く車輌の旋回時に電気式制御装置２４により制御される。
【００７５】
電気式制御装置２４には車高センサ２６ＦＬ、２６ＦＲ、２６ＲＬ、２６ＲＲより車輪１０ＦＬ〜１０ＲＲのストロークＸｉ（ｉ＝ｆｌ、ｆｒ、ｒｌ、ｒｒ）を示す信号、横加速度センサ２８より車体の横加速度Ｇｙを示す信号、ヨーレートセンサ３０より車輌のヨーレートγを示す信号、車速センサ３２より車速Ｖを示す信号、操舵角センサ３４より操舵角δを示す信号が入力される。
【００７６】
尚図には詳細に示されていないが、電気式制御装置２４は例えばＣＰＵとＲＯＭとＲＡＭと入出力ポート装置とを有し、これらが双方向性のコモンバスにより互いに接続された一般的な構成のマイクロコンピュータを含んでいる。また車高センサ２６ＦＬ〜２６ＲＲは車輪のバウンド方向を正として車輪のストロークＸｉを検出し、横加速度センサ２８及び操舵角センサ３４は車輌の左旋回方向を正としてそれぞれ横加速度及び操舵角を検出する。
【００７７】
電気式制御装置２４は、それぞれ図８（Ａ）及び（Ｂ）に示された前輪側及び後輪側の車輌モデルに基づきショックアブソーバ２２ＦＬ〜２２ＲＲの減衰係数を制御する。特にこの実施形態の電気式制御装置２４は、後述の如く図２及び図３に示されたフローチャートに従って横加速度Ｇｙに基づき車輌が過渡旋回状態にあるか否かを判別し、車輌が定常旋回状態にあるときには各車輪のショックアブソーバの減衰係数Ｃｉを予め設定されたハードの減衰係数Ｃｈｉｇｈに制御し、車輌が過渡旋回状態にあっても、車体のロール量が増大する過程に於いては旋回内側のショックアブソーバの減衰係数が旋回外側の減衰係数よりも高くなるよう制御し、逆に車体のロール量が減少する過程に於いては旋回外側のショックアブソーバの減衰係数が旋回内側の減衰係数よりも高くなるよう各ショックアブソーバの減衰係数を制御し、これにより過渡旋回時に於ける車高を低下させ車体の重心を低下させる。
【００７８】
また電気式制御装置２４は、車速Ｖ及び操舵角δに基づき車輌の基準ヨーレートγｔを演算し、車輌の実際のヨーレートγと基準ヨーレートγｔとの偏差Δγを演算し、偏差Δγに基づき車輌がオーバステア状態又はアンダステア状態にあるか否かを判定し、車輌がオーバステア状態又はアンダステア状態にあるときには偏差Δγが小さくなるよう所定の距離の補正量ΔＬｆ、ΔＬｒ及び前輪側及び後輪側の各仮想のショックアブソーバの減衰係数の補正量ΔＣｇｆ、ΔＣｇｒ、ΔＣｆ、ΔＣｒを演算し、その演算結果に基づき実際の各ショックアブソーバの減衰係数を制御する。
【００７９】
次に図２及び図３に示されたフローチャートを参照して図示の第一の実施形態に於ける減衰係数の制御について説明する。尚図２及び図３に示されたフローチャートによる制御は図には示されていないイグニッションスイッチの閉成により開始され、所定の時間毎に繰返し実行される。
【００８０】
まずステップ１０に於いては各車輪のストロークＸｉを示す信号等の読み込みが行われ、ステップ２０に於いては検出されたヨーレートγよりノイズ成分を除去するためのフィルタ処理が行われることによりフィルタ処理後のヨーレートγｆが演算される。
【００８１】
ステップ３０に於いては操舵角δに基づき前輪の実舵角δfが演算され、ＨをホイールベースとしＫhをスタビリティファクタとして下記の式３０に従って目標ヨーレートγeが演算されると共に、Ｔを時定数としｓをラプラス演算子として下記の式３１に従って車速Ｖ及び操舵角δに基づく車輌の基準ヨーレートγtが演算される。尚目標ヨーレートγeは動的なヨーレートを考慮すべく車輌の横加速度Ｇyを加味して演算されてもよい。
γe＝Ｖδf／（１＋ＫhＶ２）Ｈ ……（３０）
γt＝γe／（１＋Ｔｓ） ……（３１）
【００８２】
ステップ４０に於いては下記の式３２に従ってヨーレート偏差Δγ、即ちフィルタ処理後のヨーレートγｆと基準ヨーレートγｔとの偏差が演算される。
Δγ＝γｆ−γｔ ……（３２）
【００８３】
ステップ５０に於いてはフィルタ処理後のヨーレートγｆが正の値であるか否かの判別、即ち車輌が左旋回状態にあるか否かの判別が行われ、否定判別が行われたときにはステップ７０へ進み、肯定判別が行われたときにはステップ６０へ進む。
【００８４】
ステップ６０に於いてはヨーレート偏差Δγが正の値であるか否かの判別、即ち車輌がオーバステア状態にあるか否かの判別が行われ、肯定判別が行われたときにはステップ８０へ進み、否定判別が行われたときにはステップ９０へ進む。
【００８５】
同様にステップ７０に於いてはヨーレート偏差Δγが負の値であるか否かの判別、即ち車輌がオーバステア状態にあるか否かの判別が行われ、肯定判別が行われたときにはステップ８０へ進み、否定判別が行われたときにはステップ９０へ進む。
【００８６】
ステップ８０に於いてはヨーレート偏差Δγに基づき図４の第一及び第四象限に示されたグラフに対応するマップより前輪側の車輌モデルの所定の距離の補正量ΔＬｆ、後輪側の車輌モデルの所定の距離の補正量ΔＬｒ、前輪側の仮想のショックアブソーバ１２２Ｆの減衰係数の補正量ΔＣｇｆ、前輪側の仮想のショックアブソーバ１２４Ｆの減衰係数の補正量ΔＣｆ、後輪側の仮想のショックアブソーバ１２２Ｒの減衰係数の補正量ΔＣｇｒ、後輪側の仮想のショックアブソーバ１２４Ｒの減衰係数の補正量ΔＣｒが演算される。
【００８７】
同様にステップ９０に於いてはヨーレート偏差Δγに基づき図４の第二及び第三象限に示されたグラフに対応するマップより所定の距離の補正量ΔＬｆ、ΔＬｒ及び減衰係数の補正量ΔＣｇｆ、ΔＣｆ、ΔＣｇｒ、ΔＣｒが演算される。
【００８８】
ステップ１００に於いてはＬｆｏ及びＬｒｏをそれぞれ前輪側及び後輪側の車輌モデルの基本の所定の距離とし、Ｃｇｆｏ、Ｃｇｒｏ、Ｃｆｏ、Ｃｒｏをそれぞれ前輪側の仮想のショックアブソーバ１２２Ｆ、後輪側の仮想のショックアブソーバ１２２Ｒ、前輪側の仮想のショックアブソーバ１２４Ｆ及び後輪側の仮想のショックアブソーバ１２４Ｒについて予め設定された基本の減衰係数として、下記の式３３〜３８に従って所定の距離Ｌｆ、Ｌｒ及び各仮想のショックアブソーバの減衰係数Ｃｇｆ、Ｃｇｒ、Ｃｆ、Ｃｒが演算される。
【００８９】
Ｌｆ＝Ｌｆｏ＋ΔＬｆ ……（３３）
Ｌｒ＝Ｌｒｏ＋ΔＬｒ ……（３４）
Ｃｇｆ＝Ｃｇｆｏ＋ΔＣｇｆ ……（３５）
Ｃｇｒ＝Ｃｇｒｏ＋ΔＣｇｒ ……（３６）
Ｃｆ＝Ｃｆｏ＋ΔＣｆ ……（３７）
Ｃｒ＝Ｃｒｏ＋ΔＣｒ ……（３８）
【００９０】
ステップ１１０に於いては例えば所定の距離Ｌｆ及びＬｒの前回値と今回値との偏差ΔＬｆａ及びΔＬｒａが演算されると共に、偏差ΔＬｆａ及びΔＬｒａの絶対値が基準値Ｌｏ（正の定数）を越えているときには偏差の絶対値がＬｏになるよう今回値が補正されることにより、所定の距離Ｌｆ及びＬｒの変化率が制限される処理が行われる。
【００９１】
ステップ１２０に於いては例えば減衰係数Ｃｇｆ及びＣｇｒの前回値と今回値との偏差ΔＣｇｆａ及びΔＣｇｒａが演算されると共に、偏差ΔＣｇｆａ及びΔＣｇｒａの絶対値が基準値Ｃｇｏ（正の定数）を越えているときには偏差の絶対値がＣｇｏになるよう今回値が補正されることにより、減衰係数Ｃｇｆ及びＣｇｒの変化率が制限される処理が行われる。また例えば減衰係数Ｃｆ及びＣｒの前回値と今回値との偏差ΔＣｆａ及びΔＣｒａが演算されると共に、偏差ΔＣｆａ及びΔＣｒａの絶対値が基準値Ｃｏ（正の定数）を越えているときには偏差の絶対値がＣｏになるよう今回値が補正されることにより、減衰係数Ｃｆ及びＣｒの変化率が制限される処理が行われ、しかる後るステップ６２０へ進む。
【００９２】
ステップ６２０に於いては横加速度Ｇｙ の絶対値が制御のしきい値としての基準値Ｇｙｏ（正の定数）を越えているか否かの判別、即ち車輪の旋回時に於けるショックアブソーバの減衰係数の制御が必要であるか否かの判別が行われ、肯定判別が行われたときにはステップ６４０へ進み、否定判別が行われたときにはステップ６３０へ進む。
【００９３】
ステップ６３０に於いては各車輪のショックアブソーバの減衰係数が車輌の非旋回時に於ける通常の制御ルーチンに従って設定され、しかる後ステップ７８０へ進む。尚この場合の減衰係数の制御は当技術分野に於いて公知の任意の要領にて行われてよい。
【００９４】
ステップ６４０に於いては横加速度Ｇｙ の時間微分値ΔＧｙ が演算されると共に、時間微分値ΔＧｙ の絶対値がその基準値ΔＧｙｏ（正の定数）を越えているか否かの判別、即ち車輌が過渡旋回状態にあるか否かの判別が行われ、肯定判別が行われたときにはステップ６６０へ進み、否定判別が行われたときはステップ６５０に於いて各車輪のショックアブソーバの減衰係数Ｃｉ が予め設定されたハードの減衰係数Ｃｈｉｇｈに設定された後ステップ７８０へ進む。
【００９５】
ステップ６６０に於いては各車輪のストロークＸｉ の時間微分値（ストローク速度）Ｘｉｄ（ｉ＝ｆｌ、ｆｒ、ｒｌ、ｒｒ）が演算され、ステップ６７０に於いては横加速度Ｇｙ が正であるか否かの判別、即ち車輌が左旋回状態にあるか否かの判別が行われ、肯定判別が行われたときにはステップ６８０へ進み、否定判別が行われたときにはステップ６９０へ進む。
【００９６】
ステップ６８０に於いては旋回内側前輪のストローク速度Ｘｆｉｎｄが左前輪のストローク速度Ｘｆｌｄ に設定され、旋回外側前輪のストローク速度Ｘｆｏｕｔｄ が右前輪のストローク速度Ｘｆｒｄ に設定され、旋回内側後輪のストローク速度Ｘｒｉｎｄが左後輪のストローク速度Ｘｒｌｄ に設定され、旋回外側後輪のストローク速度Ｘｒｏｕｔｄ が右後輪のストローク速度Ｘｒｒｄ に設定される。
【００９７】
同様にステップ６９０に於いては旋回内側前輪のストローク速度Ｘｆｉｎｄが右前輪のストローク速度Ｘｆｒｄ に設定され、旋回外側前輪のストローク速度Ｘｆｏｕｔｄ が左前輪のストローク速度Ｘｆｌｄ に設定され、旋回内側後輪のストローク速度Ｘｒｉｎｄが右後輪のストローク速度Ｘｒｒｄ に設定され、旋回外側後輪のストローク速度Ｘｒｏｕｔｄ が左後輪のストローク速度Ｘｒｌｄ に設定される。
【００９８】
ステップ７００に於いてはｓｉｇｎＧｙ を横加速度Ｇｙ の符号として横加速度の時間微分値ΔＧｙ とｓｉｇｎＧｙ との積が正であるか否かの判別、即ち車輌の旋回に起因する横加速度の大きさが増大過程にあり車体のロール量が増大する状況にあるか否かの判別が行われ、肯定判別が行われたときはステップ７１０に於いて旋回内側前輪、旋回外側前輪、旋回内側後輪、旋回外側後輪のショックアブソーバの減衰係数Ｃｊ （ｊ＝ｆｉｎ 、ｆｏｕｔ、ｒｉｎ 、ｒｏｕｔ）が前記式２２〜２５に従って演算され、否定判別が行われたときにはステップ７２０に於いて各ショックアブソーバの減衰係数Ｃｊ が前記式２６〜２９に従って演算される。
【００９９】
ステップ７３０に於いては左前輪のショックアブソーバの減衰係数Ｃｆｌが旋回内側前輪の減衰係数Ｃｆｉｎ に設定され、右前輪のショックアブソーバの減衰係数Ｃｆｒが旋回外側前輪の減衰係数Ｃｆｏｕｔに設定され、左後輪のショックアブソーバの減衰係数Ｃｒｌが旋回内側後輪の減衰係数Ｃｒｉｎ に設定され、右後輪のショックアブソーバの減衰係数Ｃｒｒが旋回外側後輪の減衰係数Ｃｒｏｕｔに設定される。
【０１００】
同様にステップ７４０に於いては横加速度の時間微分値ΔＧｙ とｓｉｇｎＧｙ との積が正であるか否かの判別が行われ、否定判別が行われたときはステップ７５０に於いて旋回内側前輪、旋回外側前輪、旋回内側後輪、旋回外側後輪のショックアブソーバの減衰係数Ｃｊ が前記式２２〜２５に従って演算され、肯定判別が行われたときにはステップ７６０に於いて各ショックアブソーバの減衰係数Ｃｊ が前記式２６〜２９に従って演算される。
【０１０１】
ステップ７７０に於いては左前輪のショックアブソーバの減衰係数Ｃｆｌが旋回外側前輪の減衰係数Ｃｆｏｕｔに設定され、右前輪のショックアブソーバの減衰係数Ｃｆｒが旋回内側前輪の減衰係数Ｃｆｉｎ に設定され、左後輪のショックアブソーバの減衰係数Ｃｒｌが旋回外側後輪の減衰係数Ｃｒｏｕｔに設定され、右後輪のショックアブソーバの減衰係数Ｃｒｒが旋回内側後輪の減衰係数Ｃｒｉｎ に設定される。
【０１０２】
ステップ７８０に於いては各ショックアブソーバの減衰係数がステップ６３０、６５０、７３０又は７７０に於いて設定された減衰係数になるよう制御され、しかる後ステップ１０へ戻る。
【０１０３】
かくして図示の第一の実施形態によれば、ステップ６２０に於いて車輌の旋回時に於けるショックアブソーバの減衰係数の制御が必要であるか否かの判別が行われ、ステップ６４０に於いて車輌が過渡旋回状態にあるか否かの判別が行われ、ステップ６７０に於いて車輌の旋回方向が判定され、ステップ６６０、６８０及び６９０に於いて各車輪のストローク速度が求められ、ステップ７００及び７４０に於いて車体のロール量が増大する過程にあるか否かの判別が行われ、車体のロール量が増大する過程にあるときにはステップ７１０及び７５０に於いて各ショックアブソーバの減衰係数Ｃj が式２２〜２５に従って演算され、車体のロール量が減少する過程にあるときにはステップ７２０及び７６０に於いて各ショックアブソーバの減衰係数Ｃj が式２６〜２９に従って演算される。
【０１０４】
従って図示の第一の実施形態によれば、車輌が車体のロール量が増大する過渡旋回状態にあるときには、旋回内側のショックアブソーバの減衰係数が旋回外側の減衰係数よりも高くなるよう各ショックアブソーバの減衰係数が制御され、逆に車輌が車体のロール量が減少する過渡旋回状態にあるときには、旋回外側のショックアブソーバの減衰係数が旋回内側の減衰係数よりも高くなるよう各ショックアブソーバの減衰係数が制御されるので、車高を低下させ車体の重心を低下させて過渡旋回時に於ける車輌の運動性能を向上させることができる。
【０１０５】
また図示の第一の実施形態によれば、左右前輪のショックアブソーバの減衰係数及び左右後輪のショックアブソーバの減衰係数は相互に独立して制御されるので、例えば前記式２０〜２９に於けるＷｆ 及びＷｒを適宜に設定し、補正量ΔＬｆ及びΔＬｒ、ΔＣｇｆ及びΔＣｇｒ、ΔＣｆ及びΔＣｒを演算するためのマップ（図４）を適宜に設定することにより、車輌の過渡旋回時に於ける車体の前後方向の姿勢を制御し、例えば旋回初期に於ける車体のノーズダイブを低減したり、旋回終期に於ける車体のノーズリフトを低減したりすることができる。
【０１０６】
また図示の第一の実施形態によれば、車体のロール量が増大過程又は減少過程にあるか否かの判定は車体の横加速度Ｇｙ に基づき行われるので、例えば車高センサ２６ＦＬ〜２６ＲＲにより検出される各輪のストロークＸｉ に基づき車体の実際のロール量が演算され、その実際のロール量に基づき車体ロール量が増大過程又は減少過程にあるか否かが判定される場合に比して応答性よく各ショックアブソーバの減衰係数を制御することができる。
【０１０７】
尚ステップ６２０〜７８０は第一乃至第四の実施形態に於いて共通であるので、以上の各作用効果は後述の第二乃至第四の実施形態に於いても同様に得られる。
【０１０８】
特に図示の第一の実施形態によれば、ステップ３０に於いて車輌の基準ヨーレートγｔが演算され、ステップ４０に於いて実ヨーレートγｆと基準ヨーレートγｔとの偏差Δγが演算され、ステップ５０〜７０に於いて車輌がオーバステア状態又はアンダステア状態にあるか否かの判別が行われ、車輌がオーバステア状態にあるときにはステップ８０及び１００に於いて前輪側の車輌モデルの所定の距離Ｌｆ及び前輪側の仮想のショックアブソーバの減衰係数Ｃｇｆ、Ｃｆが偏差Δγに応じて増大補正されると共に、後輪側の車輌モデルの所定の距離Ｌｒ及び後輪側の仮想のショックアブソーバの減衰係数Ｃｇｒ、Ｃｒが偏差Δγに応じて低減補正される。
【０１０９】
逆に車輌がアンダステア状態にあるときにはステップ９０及び１００に於いて前輪側の車輌モデルの所定の距離Ｌｆ及び前輪側の仮想のショックアブソーバの減衰係数Ｃｇｆ、Ｃｆが偏差Δγに応じて低減補正されると共に、後輪側の車輌モデルの所定の距離Ｌｒ及び後輪側の仮想のショックアブソーバの減衰係数Ｃｇｒ、Ｃｒが偏差Δγに応じて増大補正される。
【０１１０】
従って第一の実施形態によれば、車輌がオーバステア状態にあるときにはオーバステア状態の程度に応じて前輪側のロール剛性が増大されると共に後輪側のロール剛性が低減され、車輌がアンダステア状態にあるときにはアンダステア状態の程度に応じて前輪側のロール剛性が低減されると共に後輪側のロール剛性が増大されるので、車輌のステア変化を低減して車輌の操縦安定性を向上させることができる。
【０１１１】
また図示の第一の実施形態によれば、車輌モデルの所定の距離Ｌｆ及びＬｒに加えて仮想のショックアブソーバの減衰係数Ｃｇｆ、Ｃｆ、Ｃｇｒ、Ｃｒも偏差Δγに応じて増減補正されるので、車輌モデルの所定の距離のみが偏差Δγに応じて増減補正される場合に比して実際のショックアブソーバの減衰係数を的確に制御することができる。
【０１１２】
また図示の第一の実施形態によれば、ステップ１１０に於いて所定の距離Ｌｆ及びＬｒの変化率が制限され、またステップ１２０に於いて減衰係数Ｃｇｆ、Ｃｆ、Ｃｇｒ、Ｃｒの変化率が制限されるので、かかる変化率の制限処理が行われない場合に比してショックアブソーバの減衰力の急激な変化及びこれに起因する車輌の乗り心地性の悪化を確実に防止することができる。
【０１１３】
尚図示の第一の実施形態に於いては、車輌のヨーレートγはヨーレートセンサ３０により検出されるようになっているが、操舵輪である左右前輪の車輪速度Ｖｗｆｌ及びＶｗｆｒが検出され、Ｔｒを車輌のトレッドとして車輪速度に基づき下記の式３９に従って演算されてもよい。
γ＝（Ｖｗｆｒ−Ｖｗｆｌ）／Ｔｒ ……（３９）
【０１１４】
また図示の第一の実施形態に於いては、車輌モデルの所定の距離Ｌｆ及びＬｒに加えて仮想のショックアブソーバの減衰係数Ｃｇｆ、Ｃｆ、Ｃｇｒ、Ｃｒも偏差Δγに応じて増減補正されるようになっているが、減衰係数Ｃｇｆ、Ｃｆ、Ｃｇｒ、Ｃｒの増減補正は省略されてもよい。また図示の実施形態に於いては、同一の偏差Δγについて見て減衰係数Ｃｇｆ及びＣｇｒの増減補正量は、減衰係数Ｃｆ及びＣｒの増減補正量よりも大きく設定されているが、減衰係数Ｃｆ及びＣｒの増減補正量が減衰係数Ｃｇｆ及びＣｇｒの増減補正量よりも大きく設定されてもよく、更には減衰係数Ｃｇｆ、Ｃｇｒ及び減衰係数Ｃｆ、Ｃｒの一方の増減補正が省略されてもよい。
【０１１５】
更に図示の第一の実施形態に於いては、車輌の目標ヨーレートγｅは上記式３０に従って演算されるようになっているが、車速Ｖ及び操舵角δに基づき図５に示されたグラフに対応するマップより演算されてもよい。
【０１１６】
第二の実施形態
図９は本発明による減衰係数制御装置の第二の好ましい実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【０１１７】
図には示されていないが、この第二の実施形態の電気式制御装置２４には車輌のヨーレートγを示す信号、車速Ｖを示す信号、操舵角δを示す信号は入力されず、スロットル開度センサよりエンジンのスロットル開度Ｔｈを示す信号及びブレーキストロークセンサよりブレーキペダルの踏み込みストロークＳｂを示す信号も入力されるようになっている。
【０１１８】
またこの第二の実施形態の減衰係数制御ルーチンのステップ２１０に於いては各車輪のストロークＸｉを示す信号等の読み込みが行われ、ステップ２２０に於いては例えばスロットル開度Ｔｈの時間微分値としてスロットル開度速度Ｖｔが演算される。
【０１１９】
ステップ２３０に於いてはスロットル開度速度Ｖｔ及び車速Ｖに基づき図１０（前輪駆動車）又は図１１（後輪駆動車）に示されたグラフに対応するマップより前輪側の所定の距離Ｌｆに対する配分比Ｋａが演算される。
【０１２０】
ステップ２４０に於いてはブレーキストロークＳｂに基づき図１２に示されたグラフに対応するマップより前輪側の所定の距離Ｌｆに対する配分比Ｋｂが演算される。
【０１２１】
ステップ２５０に於いてはそれぞれ下記の式４０及び４１に従って前輪側の所定の距離Ｌｆ及び後輪側の距離Ｌｒが演算され、しかる後ステップ６２０へ進む。
Ｌｆ＝ＫａＫｂＬｆｏ ……（４０）
Ｌｒ＝（１−Ｋａ）（１−Ｋｂ）Ｌｒｏ ……（４１）
【０１２２】
かくして図示の第二の実施形態によれば、ステップ２２０に於いてスロットル開度速度Ｖｔが演算され、ステップ２３０に於いてスロットル開度速度Ｖｔ及び車速Ｖに基づき前輪側の所定の距離Ｌｆに対する配分比Ｋａが演算され、ステップ２５０に於いて配分比Ｋａに基づく前後輪の配分比にて前輪側の所定の距離Ｌｆ及び後輪側の所定の距離Ｌｒが演算される。
【０１２３】
一般に、前輪駆動車の場合には車輌の加速時に前輪の駆動力に起因して車輌のステア特性がアンダステア側へ変化し、逆に後輪駆動車の場合には後輪の駆動力に起因して車輌のステア特性がオーバステア側へ変化するが、図示の第二の実施形態によれば、車輌が前輪駆動車である場合には車輌の加速時に前輪側の所定の距離Ｌｆに対する後輪側の所定の距離Ｌｒの比が増大されるので、車輌のアンダステア側へのステア特性の変化が低減され、また車輌が後輪駆動車である場合には車輌の加速時に後輪側の所定の距離Ｌｒに対する前輪側の所定の距離Ｌｆの比が増大されるので、車輌のオーバステア側へのステア特性の変化が低減され、従って車輌の操縦安定性を向上させることができる。
【０１２４】
また図示の第二の実施形態によれば、ステップ２４０に於いてブレーキストロークＳｂに基づき前輪側の所定の距離Ｌｆに対する配分比Ｋｂが演算され、ステップ２５０に於いて配分比Ｋｂに基づく前後輪の配分比にて前輪側の所定の距離Ｌｆ及び後輪側の所定の距離Ｌｒが演算される。
【０１２５】
従って第二の実施形態によれば、車輌の制動時に於ける車輌前方への荷重移動に起因するオーバステア側へのステア特性の変化を低減し、これにより車輌の操縦安定性を向上させることができる。
【０１２６】
また図示の第二の実施形態によれば、車輌の加減速度は運転者の制動操作量であるブレーキストロークＳｂ及び運転者の加速操作量であるスロットル開度速度に基づき推定されるので、車輌の加減速度が例えば前後加速度センサにより検出される場合に比して応答性よく各車輪のショックアブソーバの減衰係数を制御することができる。
【０１２７】
尚図示の第二の実施形態に於いては、車輌のピッチング状態量としての車輌の加速度はスロットル開度速度Ｖｔに基づき推定されるようになっているが、車輌の加速度は例えば自動変速機のトルクコンバータの出力トルク等に基づき推定されてもよい。また車輌のピッチング状態量としての車輌の減速度はブレーキストロークＳｂに基づき推定されるようになっているが、車輌の減速度は例えば図には示されていないブレーキペダルの踏力やブレーキマスタシリンダ内の圧力に基づき推定されてもよい。
【０１２８】
第三の実施形態
図１３は本発明による減衰係数制御装置の第三の好ましい実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【０１２９】
またこの第三の実施形態の減衰係数制御ルーチンのステップ２６０に於いては各車輪のストロークＸｉを示す信号等の読み込みが行われ、ステップ２７０に於いては例えば操舵角δの時間微分値として操舵角速度δｄが演算されると共に、Ｋ１及びＫ２をそれぞれ正の定数として下記の式４２に従って車輌の横加加速度の大きさＪｙが演算される。
Ｊｙ＝｜Ｋ１δＶ^２＋Ｋ２δＶ^２｜ ……（４２）
【０１３０】
ステップ２８０に於いては車輌の横加加速度の大きさＪｙに基づき図１４に示されたグラフに対応するマップより前輪側の所定の距離Ｌｆ及び後輪側の所定の距離Ｌｒが演算され、しかる後ステップ６２０へ進む。
【０１３１】
かくして図示の第三の実施形態によれば、車輌の横加加速度の大きさＪｙが大きいほど大きくなるよう所定の距離Ｌｆ及びＬｒが可変設定されるので、車体のロール角の変化が生じ易いほど車輌のロール剛性を増大し、これにより車輌の過渡旋回時に於ける車体のロールの姿勢変化を車輌の旋回状況に応じて適切に抑制することができる。
【０１３２】
また図示の第三の実施形態によれば、車輌の横加加速度の大きさＪｙが極端に大きいときには、横加加速度の大きさＪｙの増大につれて所定の距離Ｌｆ及びＬｒが漸次低減されるので、運転者により非常に急激な過渡旋回操作が行われるときには車体がロールし、これにより運転者に減速するなどの適当な処置を講ずるよう注意を喚起することができる。
【０１３３】
第四の実施形態
図１５は本発明による減衰係数制御装置の第四の好ましい実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【０１３４】
図には示されていないが、この第四の実施形態の電気式制御装置２４には車輌のヨーレートγを示す信号、車速Ｖを示す信号、操舵角δを示す信号は入力されないが、対応する圧力センサにより検出されるショックアブソーバ２２ＦＬ〜２２ＲＲのシリンダ上室内の圧力Ｐｕｉ（ｉ＝ｆｌ、ｆｒ、ｒｌ、ｒｒ）及びシリンダ下室内の圧力Ｐｌｉ（ｉ＝ｆｌ、ｆｒ、ｒｌ、ｒｒ）を示す信号も入力されるようになっている。
【０１３５】
またこの第四の実施形態の減衰係数制御ルーチンのステップ３１０に於いては各車輪のストロークＸｉを示す信号等の読み込みが行われ、ステップ３２０に於いては下記の式４３乃至式４６に従ってそれぞれショックアブソーバ２２ＦＬ〜２２ＲＲのシリンダ上室内の圧力Ｐｕｉとシリンダ下室内の圧力Ｐｌｉとの差圧Ｐｄｆｉ（ｉ＝ｆｌ、ｆｒ、ｒｌ、ｒｒ）が演算される。
【０１３６】
Ｐｄｆｆｌ＝Ｐｕｆｌ−Ｐｌｆｌ ……（４３）
Ｐｄｆｆｒ＝Ｐｕｆｒ−Ｐｌｆｒ ……（４４）
Ｐｄｆｒｌ＝Ｐｕｒｌ−Ｐｌｒｌ ……（４５）
Ｐｄｆｒｒ＝Ｐｕｒｒ−Ｐｌｒｒ ……（４６）
【０１３７】
ステップ３３０に於いてはＭＡＸを（ ）内の数値の大きい方の値として下記の式４７に従って前輪側の代表差圧Ｐｄｆｆが演算され、ステップ３４０に於いては前輪側の代表差圧Ｐｄｆｆに基づき図１６に示されたグラフに対応するマップより前輪側の所定の距離Ｌｆが演算される。
Ｐｄｆｆ＝ＭＡＸ（Ｐｄｆｆｌ，Ｐｄｆｆｒ） ……（４７）
【０１３８】
ステップ３５０に於いては下記の式４８に従って後輪側の代表差圧Ｐｄｆｒが演算され、ステップ３６０に於いては後輪側の代表差圧Ｐｄｆｒに基づき図１７に示されたグラフに対応するマップより後輪側の所定の距離Ｌｒが演算され、しかる後ステップ６２０へ進む。
Ｐｄｆｒ＝ＭＡＸ（Ｐｄｆｒｌ，Ｐｄｆｒｒ） ……（４８）
【０１３９】
かくしてこの第四の実施形態によれば、ステップ３２０に於いて各ショックアブソーバ２２ＦＬ〜２２ＲＲのシリンダ上室内の圧力Ｐｕｉとシリンダ下室内の圧力Ｐｌｉとの差圧Ｐｄｆｉが演算され、ステップ３３０に於いて差圧Ｐｄｆｆｌ及びＰｄｆｆｒの大きい方の値が前輪側の代表差圧Ｐｄｆｆに設定され、ステップ３４０に於いて前輪側の代表差圧Ｐｄｆｆに基づき代表差圧が大きいほど大きくなるよう前輪側の所定の距離Ｌｆが演算される。ステップ３５０に於いて差圧Ｐｄｆｒｌ及びＰｄｆｒｒの大きい方の値が後輪側の代表差圧Ｐｄｆｒに設定され、ステップ３６０に於いて後輪側の代表差圧Ｐｄｆｒに基づき代表差圧が大きいほど大きくなるよう後輪側の所定の距離Ｌｒが演算される。
【０１４０】
従って第四の実施形態によれば、代表差圧が大きいほど、換言すれば車体と車輪との間の相対変位の速度が高いほど車輌モデルの所定の距離が増大されることによって実際のショックアブソーバの減衰係数が増大されるので、車輌の過渡旋回時に於ける外乱に起因する車体の姿勢変化を低減することができ、また車輪の接地性を向上させて車輌の過渡旋回時の運動性能を向上させることができる。
【０１４１】
第五の実施形態
図１９は本発明による減衰係数制御装置の第五の好ましい実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【０１４２】
この第五の実施形態の減衰係数制御ルーチンのステップ４１０に於いては各車輪のストロークＸｉを示す信号等の読み込みが行われ、ステップ４２０に於いては横加速度Ｇｙと車速Ｖ及びヨーレートγの積Ｖγとの偏差Ｇｙ−Ｖγとして横加速度の偏差、即ち車輌の横すべり加速度Ｖｙｄが演算され、横すべり加速度Ｖｙｄが積分されることにより車体の横すべり速度Ｖｙが演算され、車体の前後速度Ｖｘ（＝車速Ｖ）に対する車体の横すべり速度Ｖｙの比Ｖｙ／Ｖｘとして車体のスリップ角βが演算される。
【０１４３】
ステップ４３０に於いては例えば車体のスリップ角βの時間微分値として車体のスリップ角の変化率βｄが演算されると共に、車体のスリップ角β及びその変化率βｄに基づき図２０に示されたグラフに対応するマップより車輌がどの領域にあるかの判定が行われ、ステップ４４０に於いては前輪側の所定の距離Ｌｆに対する補正係数Ｋｆ及び後輪側の所定の距離Ｌｒに対する補正係数Ｋｒがそれぞれ１にセットされる。
【０１４４】
尚図２０のグラフに於いて、領域Ａは車輌の通常の安定的な走行状態を示し、領域Ｂは車輌がスピン傾向の状態にある領域を示し、領域Ｃは車輌がドリフトアウト傾向の状態にある領域を示し、領域Ｄは車輌の旋回挙動が発散する領域を示している。
【０１４５】
ステップ４５０に於いては車輌が領域Ｂにあるか否かの判別が行われ、否定判別が行われたときにはステップ４７０へ進み、肯定判別が行われたときにはステップ４６０に於いて車体のスリップ角βの絶対値に基づき図２１に示されたグラフに対応するマップより補正係数Ｋｆが演算される。
【０１４６】
ステップ４７０に於いては車輌が領域Ｃにあるか否かの判別が行われ、否定判別が行われたときにはステップ４９０へ進み、肯定判別が行われたときはにステップ４８０に於いて車体のスリップ角βの絶対値に基づき図２２に示されたグラフに対応するマップより補正係数Ｋｒが演算される。
【０１４７】
ステップ４９０に於いては車輌が領域Ｄにあるか否かの判別が行われ、否定判別が行われたときにはステップ５１０へ進み、肯定判別が行われたときにはステップ５００に於いて車体のスリップ角βの絶対値に基づき図２３に示されたグラフに対応するマップより補正係数Ｋf及びＫrが演算される。
【０１４８】
ステップ５１０に於いては前輪側の所定の距離Ｌｆ及び後輪側の所定の距離Ｌｒがそれぞれ下記の式４９及び５０に従って演算され、ステップ５２０に於いては第一の実施形態に於けるステップ１１０の場合と同一の要領にて所定の距離Ｌｆ及びＬｒの変化率の制限処理が行われ、しかる後ステップ６２０へ進む。
Ｌｆ＝ＫｆＬｆｏ ……（４９）
Ｌｒ＝ＫｒＬｒｏ ……（５０）
【０１４９】
かくして図示の第五の実施形態によれば、ステップ４２０に於いて車体のスリップ角βが演算され、ステップ４３０に於いて車体のスリップ角の変化率βｄが演算されると共に、車体のスリップ角β及びその変化率βｄに車輌が旋回挙動のどの領域にあるかの判定が行われ、ステップ４５０に於いて車輌が領域Ｂ、即ちスピン傾向の状態にあると判定されたときには、ステップ４６０に於いて車体のスリップ角βの絶対値が大きいほど大きくなるよう前輪の車輌モデルの所定の距離Ｌｆが増大され、これにより前輪側の車輌のロール剛性が増大される。
【０１５０】
またステップ４７０に於いて車輌が領域Ｃ、即ちドリフトアウト傾向の状態にあると判定されると、ステップ４８０に於いて車体のスリップ角βの絶対値が大きいほど大きくなるよう後輪の車輌モデルの所定の距離Ｌｒが増大され、これにより後輪側の車輌のロール剛性が増大される。
【０１５１】
更にステップ４９０に於いて車輌が領域Ｄ、即ち車輌の旋回挙動が発散する領域にあると判定されると、ステップ５００に於いて車体のスリップ角βの絶対値が大きいほど大きくなるよう前輪及び後輪の車輌モデルの所定の距離Ｌｆ及びＬｒが増大され、これにより車輌全体のロール剛性が増大される。
【０１５２】
従って第五の実施形態によれば、車輌の過渡旋回時に於ける車輌の旋回挙動を判定し、その判定結果に応じて実際のショックアブソーバの減衰係数を適切に制御し、これにより車輌のスピンやドリフトアウトの如き旋回挙動の悪化を効果的に抑制することができる。
【０１５３】
また第五の実施形態によれば、ステップ５２０に於いて所定の距離Ｌｆ及びＬｒの変化率が制限されるので、かかる変化率の制限処理が行われない場合に比してショックアブソーバの減衰力の急激な変化及びこれに起因する車輌の乗り心地性の悪化を確実に防止することができる。
【０１５４】
以上に於ては本発明を特定の実施形態について詳細に説明したが、本発明は上述の実施形態に限定されるものではなく、本発明の範囲内にて他の種々の実施形態が可能であることは当業者にとって明らかであろう。
【０１５５】
例えば上述の各実施形態に於いては、車体ロール量の増大過程に於いては旋回内側のショックアブソーバの減衰係数が旋回外側のショックアブソーバの減衰係数よりも相対的に高く制御され、逆に車体ロール量の減少過程に於いては旋回外側のショックアブソーバの減衰係数が旋回内側のショックアブソーバの減衰係数よりも高く制御されるようになっているが、一般に車体ロール量の減少過程（旋回終期）に於いて車輌の挙動が不安定になる虞れは車体ロール量の増大過程（旋回初期）に比して低いので、車体ロール量の減少過程に於いて旋回外側のショックアブソーバの減衰係数を旋回内側のショックアブソーバの減衰係数よりも高くする制御が省略されてもよい。
【０１５６】
具体的にはステップ７２０及び７６０に於ける減衰係数Ｃｊ の演算が省略され、その代わりに各ショックアブソーバの減衰係数Ｃｉ が例えばステップ６５０の場合と同様ハードの減衰係数Ｃｈｉｇｈに設定され、しかる後ステップ７８０へ進むよう修正されてもよい。
【０１５７】
また上述の各実施形態に於いては、車輌の旋回方向の判定は車体の横加速度Ｇｙの符号に基づき判定されるようになっているが、例えばＫｈをスタビリティファクタとし、Ｒｇをステアリングギヤ比とし、Ｈをホイールベースとして、図１に示された操舵角センサ２８により検出される操舵角δ及び図１には示されていない車速センサにより検出される車速Ｖに基づき、下記の式５１に基づき車輌の横加速度Ｇｙｓが推定され、その推定された横加速度に基づき行われてもよい。
Ｇｙｓ＝Ｖ^２δ／［（１＋Ｋｈ Ｖ^２）Ｒｇ Ｈ］ ……（５１）
【０１５８】
同様に、車速及び操舵角に基づき車輌のヨーレートγｈが推定され、その符号に基づき車輌の旋回方向が判定されてもよい。尚かくして車体の推定横加速度Ｇｙｓ又は推定ヨーレートγｈに基づき車輌の旋回方向が判定される場合には、例えばカウンタステアの場合の如く判定される車輌の旋回方向が車輌の実際の旋回方向とは逆になる場合があるので、（１）車体の推定横加速度Ｇｙｓ又は推定ヨーレートγｈに基づく旋回方向の判定と、（２）車体の実際の横加速度Ｇｙ又は左右の操舵輪の車輪速度に基づき推定されるヨーレートγｈ又は左右のショックアブソーバの内圧の差又は左右の車輪のストローク速度の差又は左右のばね上速度の差に基づく旋回方向の判定とが行われ、両者の判定が異なるときには後者の判定が採用されるよう修正されてもよい。
【０１５９】
また上述の各実施形態に於いては、車体の横加速度Ｇｙの時間微分値ΔＧｙの符号に基づき車体ロール量が増大過程又は減少過程にあるか否かの判定が行われるようになっているが、この判定は例えば上記の式５１に従って演算される推定横加速度Ｇｙｓに基づき行われてもよく、また車高センサ２６ＦＬ〜２６ＲＲにより検出されるストロークＸｉに基づき演算される車体のロールレートの符号に基づき行われてもよい。またこの場合ロールレートは図１には示されていないロールレートセンサにより検出されてもよい。
【０１６０】
また上述の各実施形態に於いては、各車輪のストローク速度Ｘｉｄは車高センサ２６ＦＬ〜２６ＲＲの検出結果に基づき演算されるようになっているが、各車輪のストローク速度は車体に設けられた上下加速度センサ（図示せず）により検出される車体の上下加速度Ｇｂｉに基づきオブザーバにより推定され、車高センサが省略されてもよい。
【０１６１】
また上述の第二乃至第五の実施形態に於いては、所定の距離Ｌｆ、Ｌｒのみが可変設定されるようになっているが、第一の実施形態の場合と同様、これらの実施形態に於いても所定の距離Ｌｆ、Ｌｒと共に仮想のショックアブソーバの減衰係数Ｃｇｆ、Ｃｇｒ、Ｃｆ、Ｃｒが車輌の状態量に応じて可変設定されてもよい。
【０１６２】
また上述の第一の実施形態に於いては、所定の距離Ｌｆ、Ｌｒ及び仮想のショックアブソーバの減衰係数Ｃｇｆ、Ｃｇｒ、Ｃｆ、Ｃｒの変化率が制限され、第五の実施形態に於いては所定の距離Ｌｆ、Ｌｒの変化率が制限されるようになっているが、これらの変化率制限処理は省略されてもよく、また第二乃至第四の実施形態に於いても所定の距離Ｌｆ、Ｌｒの変化率が制限されるよう修正されてもよい。
【０１６３】
更に上述の第二の実施形態に於いては、スロットル開度速度Ｖｔ及びブレーキストロークＳｂの両者に基づき所定の距離Ｌｆ、Ｌｒ及びそれらの比が変更されるようになっているが、所定の距離Ｌｆ、Ｌｒ及びそれらの比はスロットル開度速度Ｖｔ及びブレーキストロークＳｂの一方のみに基づき変更されるよう修正されてもよい。
【０１６４】
【発明の効果】
以上の説明より明らかである如く、本発明の請求項１の構成によれば、車輌の過渡旋回時に第一の仮想のショックアブソーバによってばね上のロール変位が抑制され、第二の仮想のショックアブソーバによってばね上の旋回内輪側のリフトが抑制されるので、ばね上のロール変位を低減しばね上の重心を低下させて車輌の過渡旋回時の運動性能を向上させることができると共に、車輌の状態量に応じて車輌モデルの所定の距離が可変設定されるので、実際のショックアブソーバの減衰係数を車輌の状態量に応じて適切に制御し、これにより車輌の状態量の如何に拘わらず車輌モデルの所定の距離が一定である場合に比して、車輌の過渡旋回時に於ける車輌のステア特性の変化やばね上の姿勢変化の抑制を車輌の走行状態に応じて適切に行うことができる。
【０１６５】
また請求項２の構成によれば、車輌モデルは前輪の車輌モデルと後輪の車輌モデルとよりなるので、前輪側及び後輪側の実際のショックアブソーバの減衰係数を車輌の状態量に応じて適切に制御し、これにより車輌モデルが一つである場合に比して、車輌の過渡旋回時に於ける車輌のステア特性の変化やばね上の姿勢変化の抑制を車輌の走行状態に応じて適切に制御することができる。
【０１６６】
また請求項３の構成によれば、車輌の旋回挙動に応じて車輌モデルの所定の距離が可変設定されるので、実際のショックアブソーバの減衰係数を車輌の旋回挙動に応じて適切に制御し、これにより車輌の旋回挙動の如何に拘わらず車輌モデルの所定の距離が一定である場合に比して、車輌の過渡旋回時に於ける車輌のステア特性の変化を適切に抑制することができる。
【０１６７】
また請求項４の構成によれば、車輌の横加加速度に応じて車輌モデルの所定の距離が可変設定されるので、実際のショックアブソーバの減衰係数を車輌の横加加速度に応じて適切に制御し、これにより車輌の横加加速度の如何に拘わらず車輌モデルの所定の距離が一定である場合に比して、車輌の過渡旋回時に於けるばね上のロール姿勢変化を適切に抑制することができる。
【０１６８】
また請求項５の構成によれば、実際のショックアブソーバの二つのシリンダ室内圧力の差圧に応じて車輌モデルの所定の距離が可変設定されるので、実際のショックアブソーバの減衰係数を二つのシリンダ室内圧力の差圧に応じて適切に制御し、これによりばね上の姿勢変化を適切に制御することができる。
【０１６９】
また請求項６の構成によれば、車輌の旋回挙動に応じて二つの車輌モデルの所定の距離の比が可変設定されるので、実際のショックアブソーバの減衰係数が車輌の旋回挙動に応じて適切に制御され、これにより車輌の旋回挙動の如何に拘わらず二つの車輌モデルの所定の距離の比が一定である場合に比して、車輌の過渡旋回時に於けるばね上のピッチ姿勢変化及びこれに起因するステア特性の変化を適切に抑制することができる。
【図面の簡単な説明】
【図１】本発明による減衰係数制御装置の第一の好ましい実施形態を示す概略構成図である。
【図２】第一の実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【図３】第一の実施形態に於ける減衰係数制御ルーチンの後半を示すフローチャートである。
【図４】ヨーレート偏差Δγと所定の距離の補正量ΔＬｆ、ΔＬｒ及び減衰係数の補正量ΔＣｇｆ、ΔＣｇｒ、ΔＣｆ、ΔＣｒとの関係を示すグラフである。
【図５】操舵角δ及び車速Ｖと車輌の目標ヨーレートγｅとの関係を示すグラフである。
【図６】実際の車輌の二輪モデルを示す説明図である。
【図７】車輌の旋回内側に仮想のショックアブソーバが配設された仮想モデルを示す説明図である。
【図８】車輌の旋回内側に仮想のショックアブソーバが配設された前輪側及び後輪側の仮想モデルを示す説明図である。
【図９】第二の実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【図１０】前輪駆動車についてスロットル開度速度Ｖｔと前輪側の所定の距離Ｌｆに対する配分比Ｋａとの関係を示すグラフである。
【図１１】後輪駆動車についてスロットル開度速度Ｖｔと前輪側の所定の距離Ｌｆに対する配分比Ｋａとの関係を示すグラフである。
【図１２】通常の車輌についてブレーキストロークＳｂと前輪側の所定の距離Ｌｆに対する配分比Ｋｂとの関係を示すグラフである。
【図１３】第三の実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【図１４】車輌の横加加速度の大きさＪｙと所定の距離Ｌｆ及びＬｒとの関係を示すグラフである。
【図１５】第四の実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【図１６】前輪側の代表差圧Ｐｄｆｆと前輪側の所定の距離Ｌｆとの関係を示すグラフである。
【図１７】後輪側の代表差圧Ｐｄｆｒと後輪側の所定の距離Ｌｒとの関係を示すグラフである。
【図１８】各輪のストローク速度Ｘｉｄと差圧Ｐｄｆｉとの関係を示すグラフである。
【図１９】第五の実施形態に於ける減衰係数制御ルーチンの前半を示すフローチャートである。
【図２０】車輌のスリップ角β及びその微分値βｄと各領域との関係を示すグラフである。
【図２１】車輌のスリップ角βの絶対値と前輪側の所定の距離Ｌｆに対する補正係数Ｋｆとの関係を示すグラフである。
【図２２】車輌のスリップ角βの絶対値と後輪側の所定の距離Ｌｒに対する補正係数Ｋｒとの関係を示すグラフである。
【図２３】車輌のスリップ角βの絶対値と補正係数Ｋｆ及びＫｒとの関係を示すグラフである。
【符号の説明】
１４…ステアリングホイール
１６…パワーステアリング装置
２０…車体
２４…電気式制御装置
２６ＦＬ〜２６ＲＲ…車高センサ
２８…横加速度センサ
３０…ヨーレートセンサ
３２…車速センサ
３４…操舵角センサ
１１０…車体
１１２Ｌ、１１２Ｒ…車輪
１１４Ｌ、１１４Ｒ…サスペンションスプリング
１１６Ｌ、１１６Ｒ…ショックアブソーバ
１２２、１２４…ショックアブソーバ[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a damping coefficient control apparatus for a vehicle such as an automobile, and more particularly to a damping coefficient control apparatus improved to improve the motion performance of a vehicle during a transient turn.
[0002]
[Prior art]
As one example of a damping coefficient control device for a vehicle such as an automobile provided with a shock absorber having a variable damping coefficient corresponding to each wheel, for example, Japanese Patent Application No. 10-92675 prior to the publication of the application of the present applicant. The specification and drawings include means for detecting vehicle turning information, means for determining a change in the body roll amount, and in the process of increasing the body roll amount, the damping coefficient of the shock absorber on the inside of the turn is set to the shock on the outside of the turn. There is described a vehicle damping coefficient control device characterized by having a means for controlling the damping coefficient relatively higher than that of the absorber.
[0003]
According to the damping coefficient control device according to the previous proposal, in the process of increasing the body roll amount, the damping coefficient of the shock absorber on the inside of the turn is controlled to be relatively higher than the damping coefficient of the shock absorber on the outside of the turn, As a result, the damping force of the shock absorber on the inside of the turn acting downward is controlled to be relatively higher than the damping force of the shock absorber on the outside turning, so that the downward force acting on the vehicle as a whole increases. As a result, the vehicle height can be reduced and the motion performance during the transient turning of the vehicle can be improved.
[0004]
[Problems to be solved by the invention]
In the damping coefficient control device according to the above proposal, the vehicle body is located at a virtual position spaced laterally from the vehicle body by a predetermined distance toward the side estimated to lift with respect to the center of gravity of the vehicle body based on the turning state of the vehicle. With a virtual swing center of,Around the virtual swing centerA damping force that suppresses the roll displacement of the vehicle body is generatedFirst virtual shock absorber and virtualBelow the swing center ofpositionBetween the virtual wheel and the vehicle bodyUp and down directionGenerate damping forceBased on the vehicle model having the second virtual shock absorber, the damping coefficient of the actual shock absorber with variable damping coefficient provided corresponding to each wheel is controlled.
[0005]
In general, the turning behavior of the vehicle and the posture of the vehicle body change depending on the turning state of the vehicle. For example, when the vehicle is turning, the vehicle's steering characteristics may change to the understeer side or oversteer side due to the vehicle speed, acceleration / deceleration, steering angle, road surface conditions, etc., and acceleration / deceleration will occur if the driver performs acceleration / deceleration operations. The steering characteristic changes due to the load movement in the vehicle front-rear direction due to. Further, the posture of the vehicle body at the time of turning changes due to acceleration / deceleration and steering operation.
[0006]
However, in the damping coefficient control device according to the above proposal, since the predetermined distance of the vehicle model is constant, a situation in which a change in the steering characteristic or a change in the posture of the vehicle body occurs when the vehicle turns is caused. However, there is a problem that the actual damping coefficient of the shock absorber cannot be appropriately controlled according to the traveling state of the vehicle.
[0007]
The present invention provides a virtual swing center aroundA damping force that suppresses the roll displacement of the vehicle body is generatedFirst virtual shock absorber and virtualBelow the swing center ofpositionBetween the virtual wheel and the vehicle bodyUp and down directionGenerate damping forceIn view of the above-described problems in the damping coefficient control device according to the previous proposal, which is configured to control the damping coefficient of the actual shock absorber based on the vehicle model having the second virtual shock absorber. The main problem of the present invention is to appropriately improve the performance of the vehicle during the transient turning of the vehicle by variably setting a predetermined distance of the vehicle model in accordance with the vehicle traveling state. It is to let you.
[0008]
[Means for Solving the Problems]
According to the present invention, the main problem described above is that the vehicle is turned in a vehicle damping coefficient control device provided with an actual shock absorber having a variable damping coefficient corresponding to each wheel. A means for detecting a state, a means for detecting a state quantity of the vehicle, and a side that is presumed to lift with respect to the center of gravity on the spring based on the turning state of the vehicle; Having a virtual swing center on the spring at the virtual position placed,Around the virtual swing centerGenerates damping force that suppresses roll displacement on the springFirst virtual shock absorber and said virtualBelow the swing center ofpositionBetween the virtual wheel to run and the sprungUp and down directionGenerate damping forceA vehicle model having a second virtual shock absorber, distance setting means for variably setting the predetermined distance according to a state quantity of the vehicle, and the actual model based on at least the virtual damping coefficient and the predetermined distance. This is achieved by a vehicle damping coefficient control device comprising means for calculating a target damping coefficient of a shock absorber and means for controlling the damping coefficient of the actual shock absorber based on the target damping coefficient.
[0009]
According to the configuration of the first aspect, the virtual position on the spring is set at a virtual position spaced laterally from the spring by a predetermined distance toward the side estimated to lift with respect to the center of gravity on the spring based on the turning state of the vehicle. With a rocking center of,Around the virtual swing centerGenerates a damping force that suppresses roll displacement on the springFirst virtual shock absorber and virtualBelow the swing center ofpositionBetween the virtual wheel and the sprungUp and down directionGenerate damping forceBecause it has a vehicle model with a second virtual shock absorber thatThe first virtual shock absorber suppresses the roll displacement on the spring,Swivel on a spring by a second virtual shock absorberInsideWheel side lift is suppressed, whichRoll displacement on the spring is reducedThe center of gravity on the spring is loweredThis improves the performance of the vehicle during transient turningAt the same time, since the predetermined distance of the vehicle model is variably set according to the vehicle state quantity, the actual shock absorber attenuation coefficient is appropriately controlled according to the vehicle state quantity. Regardless of whether the vehicle model has a fixed distance, the change in the vehicle steering characteristics and the change in the posture on the spring during the transient turning of the vehicle are more appropriately controlled according to the running state of the vehicle. Done.
[0010]
According to the present invention, in order to effectively achieve the main problems described above, in the configuration of claim 1, the vehicle model is configured to include a front-wheel vehicle model and a rear-wheel vehicle model. (Structure of claim 2).
[0011]
According to the configuration of the second aspect, the vehicle model is composed of the front wheel vehicle model and the rear wheel vehicle model, so that the actual shock absorber attenuation coefficient on the front wheel side and the rear wheel side depends on the vehicle state quantity It is appropriately controlled, and as a result, the change in the steering characteristics of the vehicle and the change in the posture on the spring during the transient turning of the vehicle are more appropriate depending on the running state of the vehicle. To be controlled.
[0012]
According to the present invention, in order to effectively achieve the main problem described above, in the configuration of claim 1 or 2, the means for detecting the state quantity of the vehicle detects the turning behavior of the vehicle. The distance setting means is configured to variably set the predetermined distance according to the turning behavior of the vehicle (configuration of claim 3).
[0013]
According to the configuration of the third aspect, since the predetermined distance of the vehicle model is variably set according to the turning behavior of the vehicle, the actual shock absorber attenuation coefficient is appropriately controlled according to the turning behavior of the vehicle, As a result, the change in the steering characteristic of the vehicle during the transient turning of the vehicle is appropriately suppressed as compared with the case where the predetermined distance of the vehicle model is constant regardless of the turning behavior of the vehicle.
[0014]
According to the present invention, in order to effectively achieve the above main problem, in the configuration of claim 1 or 2, the means for detecting the state quantity of the vehicle detects a lateral jerk of the vehicle. The distance setting means is configured to variably set the predetermined distance according to a lateral jerk of the vehicle.
[0015]
According to the configuration of the fourth aspect, since the predetermined distance of the vehicle model is variably set according to the lateral jerk of the vehicle, the actual shock absorber attenuation coefficient is appropriately controlled according to the lateral jerk of the vehicle, As a result, as compared with the case where the predetermined distance of the vehicle model is constant regardless of the lateral jerk of the vehicle, a change in the roll posture on the spring during the transient turning of the vehicle is appropriately suppressed.
[0016]
According to the present invention, in order to effectively achieve the main problem described above, in the configuration of claim 1 or 2, the means for detecting the state quantity of the vehicle includes two of the actual shock absorbers. The differential pressure of the cylinder chamber pressure is detected, and the distance setting means is configured to variably set the predetermined distance in accordance with the differential pressure (configuration of claim 5).
[0017]
According to the configuration of the fifth aspect, since the predetermined distance of the vehicle model is variably set according to the differential pressure between the two cylinder chamber pressures of the actual shock absorber, the actual shock absorber has a damping coefficient of two cylinders. Appropriate control is performed according to the differential pressure of the indoor pressure, and thereby the posture change on the spring is appropriately controlled.
[0018]
According to the present invention, in order to effectively achieve the main problem described above, in the configuration of claim 2, the means for detecting the state quantity of the vehicle detects the turning behavior of the vehicle, and The distance setting means is configured to variably set the ratio of the predetermined distance between the vehicle model of the front wheel and the vehicle model of the rear wheel in accordance with the turning behavior of the vehicle (configuration of claim 6).
[0019]
According to the configuration of the sixth aspect, since the ratio of the predetermined distance between the two vehicle models is variably set according to the turning behavior of the vehicle, the actual shock absorber attenuation coefficient is appropriately set according to the turning behavior of the vehicle. This is compared to the case where the ratio of the predetermined distance between the two vehicle models is constant irrespective of the turning behavior of the vehicle, and the change in the pitch posture on the spring during the transient turning of the vehicle. A change in the steering characteristic due to the is appropriately suppressed.
[0020]
[Preferred embodiment of the problem solving means]
As shown in FIG. 6, in a two-wheel model of an actual vehicle, a vehicle body 110 is supported by left and right wheels 112L and 112R, and suspension springs 114L and 114R and a shock absorber 116L are interposed between the vehicle body 110 and the wheels 112L and 112R. And 116R are shown as being disposed.
[0021]
In the actual vehicle model shown in FIG. 6, for example, if the vehicle turns left and the rightward inertial force acts on the vehicle body 110, the roll moment Mroll outwardly turns acts on the vehicle body. The roll moment is carried by the spring forces Fsl and Fsr of the left and right suspension springs 114L and 114R, and the damping forces Fal and Far of the left and right shock absorbers 116L and 116R. The vehicle body 110 rolls outwardly until the moment in the roll restraining direction due to is equal to the roll moment Mroll.
[0022]
In this case, the increase amount of the spring force Fsl of the suspension spring 114L and the decrease amount of the spring force Fsr of the suspension spring 114R are substantially equal to each other, and in the conventional vehicle, the damping coefficients of the left and right shock absorbers at the time of turning are mutually Since they are controlled to be equal, the damping forces Fal and Far of the left and right shock absorbers are also substantially equal to each other, so that the height of the vehicle's center of gravity 118 does not change substantially.
[0023]
On the other hand, as shown in FIG. 7, only suspension springs 114L and 114R are disposed between the vehicle body 110 and the left and right wheels 112L and 112R, so that the vehicle is turned inside the vehicle.Below the spaced virtual position 118 'PlacedVirtual wheel 120 andBody 110WithOne shock absorber 122 for generating a vertical damping force in between,Around the virtual position 118 'Suppress roll displacement of the car bodyGenerate damping forceConsidering a virtual model in which one shock absorber 124 is arranged, the roll moment Mroll is carried by the damping force Fas of the shock absorber 122 and the spring forces Fsl and Fsr of the left and right suspension springs 114L and 114R. The height of the center of gravity 118 is reduced by reducing the amount of increase in the vehicle height on the turning inner wheel side as compared with.
[0024]
Therefore, if the control of the virtual model as shown in FIG. 7 can be achieved in the two-wheel model of the actual vehicle shown in FIG. 6, the height of the center of gravity 118 of the vehicle is lowered in the process of increasing the body roll amount. Thus, it is possible to improve the motion performance in the initial turning of the vehicle.
[0025]
As shown in FIG. 7, the spring constants of the left and right suspension springs 114L and 114R are K, the damping coefficient of the shock absorber on the turning outer ring side is Cout, the stroke of the turning outer ring is Xout, and the shock on the turning inner ring side is The damping coefficient of the absorber 114L is Cin, the stroke of the turning inner wheel is Xin, the tread of the vehicle is W, and the distance between the center of gravity 118 of the vehicle and the shock absorber 122 isVehicle lateralThe distance is L, and the damping coefficients of the shock absorbers 122 and 124 are Cg and C, respectively.
[0026]
The mass of the vehicle body 110And roll moment of inertia M and I, respectively.Where the vertical acceleration and roll angular velocity of the vehicle body are Xbdd and θdd, respectively, and the stroke speeds of the outer turning wheel and the inner turning wheel are Xoutd and Xind, respectively, and the balance of vertical forces in the virtual model shown in FIG. From the balance of forces around the center of gravity 118, the following formulas 1 and 2 are established, respectively.
[0027]
[Expression 1]

[0028]
Assuming that Cn = WC / 2 as a parameter for attenuating the roll motion of the vehicle body, Equation 2 is expressed as Equation 3 below.
[0029]
[Expression 2]

[0030]
Further, the following equations 4 and 5 are established from the balance of forces in the vertical direction and the balance of forces around the center of gravity 118 in the two-wheel model of the actual vehicle shown in FIG.
[0031]
[Equation 3]

[0032]
From the above formulas 1 and 4, the following formula 6 is established.
[0033]
[Expression 4]

[0034]
If Cm = Cn / L, the following formula 7 is established from the above formulas 3 and 5.
[0035]
[Equation 5]

[0036]
Here, when the vertical force generated by the shock absorber 122 in the virtual model shown in FIG., 7 and the following formula8, the following formulas 9 to 11 are established.
[0037]
[Formula 6]

[0038]
[Expression 7]

[0039]
From equation 9 + equation 11, the damping coefficient Cin of the shock absorber for the inner turning wheel can be obtained as follows.
[0040]
[Equation 8]

[0041]
Further, by substituting Equation 12 into Equation 9, the damping coefficient Cout of the shock absorber for the outer turning wheel can be obtained as follows.
[0042]
[Equation 9]

[0043]
Further, the above equations 12 and 13 are rearranged so that the damping coefficients Cin and Cout of the shock absorbers of the turning inner wheel and the turning outer wheel are expressed by the following equations 14 and 15, respectively.
[0044]
[Expression 10]

[0045]
The damping forces generated by the shock absorbers on the turning inner wheel side and the turning outer wheel side can be obtained by the following equations 16 and 17, respectively.
[0046]
## EQU11 ##

[0047]
Further, based on the same concept, in the process of decreasing the vehicle body roll amount, the shock absorbers on the inner and outer wheels are turned on the basis of a virtual model in which virtual shock absorbers 122 and 124 are disposed outside the vehicle. By controlling the damping coefficients Cin and Cout of the vehicle as shown in the following equations 18 and 19, respectively, the height of the center of gravity 118 of the vehicle can be reduced, and the motion performance at the end of turning of the vehicle can be improved.
[0048]
[Expression 12]

[0049]
Therefore, according to one preferred aspect of the present invention, in the structure of claim 1, the damping coefficient Cin of the shock absorber on the inside of the turn and the damping coefficient of the shock absorber on the outside of the turn in the process of increasing the vehicle body roll amount. Cout is configured so as to be calculated in accordance with the above formulas 14 and 15, respectively (preferred mode 1).
[0050]
According to another preferred embodiment of the present invention, in the construction of claim 1, the damping coefficient of the shock absorber on the outer side of the turn is the damping factor of the shock absorber on the inner side of the turning in the body roll amount decreasing process. It is configured to be controlled higher than the coefficient (preferred aspect 2).
[0051]
According to another preferred embodiment of the present invention, in the configuration of the preferred embodiment 2, the damping coefficient Cin of the shock absorber on the inside of the turn and the shock absorber on the outside of the turn in the process of reducing the roll amount of the vehicle body. The attenuation coefficient Cout is configured so as to be calculated according to the above equations 18 and 19, respectively (preferred aspect 3).
[0052]
According to another preferred aspect of the present invention, in accordance with the structure of claim 2, the vehicle model is composed of a vehicle model on the front wheel side and a vehicle model on the rear wheel side. (Preferred aspect 4).
[0053]
Further, as shown in FIG. 8, L, W, T, Cg, and C for the front-wheel and rear-wheel vehicle models are respectively expressed as Lf and Lr, Wf and Wr, Tf and Tr, Cgf and Cgr, Cf, and Cf. Cr, the stroke speeds of the turning inner front wheel and the turning outer front wheel are Xfind and Xfoutd, respectively, the stroke speeds of the turning inner rear wheel and the turning outer rear wheel are Xrind and Xrouted, respectively, and Tf and Tr are respectively expressed by the following equations 20 and As the values represented by 21, the damping coefficient Cfin of the shock absorber for the inner front wheel and the shock coefficient Cfout of the shock absorber for the outer front wheel are calculated according to the following equations 22 and 23, respectively, in the process of increasing the vehicle body roll amount. The damping coefficient Cr of the shock absorber for the rear wheel inside the turn It is preferable that the damping coefficient Crout of the shock absorber for in and the outer wheel on the outside of the turn is calculated according to the following equations 24 and 25, respectively.
[0054]
[Formula 13]

[0055]
[Expression 14]

[0056]
Therefore, according to another preferred embodiment of the present invention, in the configuration of the preferred embodiment 4, in the process of increasing the vehicle body roll amount, the damping coefficient Cfin of the shock absorber for the turning inner front wheel and the shock for the turning outer front wheel The damping coefficient Cfout of the absorber is calculated according to the above formulas 22 and 23, respectively. The damping coefficient Crin of the shock absorber for the inner rear wheel and the damping coefficient Crout of the shock absorber for the rear rear wheel are calculated according to the above formulas 24 and 25, respectively. (Preferred aspect 5).
[0057]
Similarly, according to another preferred embodiment of the present invention, in the configuration of the preferred embodiment 3, the vehicle model is configured to include a vehicle model on the front wheel side and a vehicle model on the rear wheel side (preferred embodiment). 6).
[0058]
According to another preferred aspect of the present invention, in the configuration of the preferred aspect 6 described above, the damping coefficient Cfin of the shock absorber for the turning inner front wheel and the shock of the turning outer front wheel in the process of decreasing the vehicle body roll amount. The absorber damping coefficient Cfout is calculated in accordance with the following equations 26 and 27, respectively. (Preferred aspect 7).
[0059]
[Expression 15]

[0060]
According to another preferred aspect of the present invention, in the configuration of claim 3, the means for detecting the state quantity of the vehicle calculates the deviation between the reference yaw rate of the vehicle and the actual yaw rate of the vehicle. (Preferred aspect 8).
[0061]
According to another preferred aspect of the present invention, in the configuration of claim 3, the means for detecting the state quantity of the vehicle detects the relationship between the slip angle of the vehicle and the rate of change thereof as the turning behavior of the vehicle. (Preferred embodiment 9).
[0062]
According to another preferred aspect of the present invention, in the configuration of claim 4, the means for detecting the state quantity of the vehicle estimates the lateral jerk of the vehicle based on the vehicle speed, the steering angle, and the steering angular velocity. Constructed (preferred embodiment 10).
[0063]
According to another preferred aspect of the present invention, in the configuration of claim 5, according to the configuration of claim 2, the distance setting means adjusts the front wheel according to the value of the larger differential pressure between the left and right front wheels. The predetermined distance of the vehicle model is variably set, and the predetermined distance of the vehicle model of the rear wheel is variably set according to the value of the larger differential pressure between the left and right rear wheels (preferable aspect 11).
[0064]
According to another preferred aspect of the present invention, in the configuration of claim 6, when the turning behavior of the vehicle is a behavior of changing the steering characteristic toward the oversteer side, the roll rigidity on the front wheel side is increased or A ratio of a predetermined distance between the vehicle model of the front wheels and the vehicle model of the rear wheels is variably set so as to reduce the roll rigidity on the rear wheel side (preferred aspect 12).
[0065]
According to another preferred embodiment of the present invention, in the configuration of claim 6, when the turning behavior of the vehicle is the behavior of the change of the steer characteristic toward the understeer side, the roll rigidity on the front wheel side is reduced or A ratio of a predetermined distance between the vehicle model of the front wheel and the vehicle model of the rear wheel is variably set so as to increase the roll rigidity on the rear wheel side (preferred aspect 13).
[0066]
According to another preferred aspect of the present invention, in the configuration of claim 6, the means for detecting the vehicle state quantity is configured to detect a pitch motion state quantity on the spring (preferred aspect). 14).
[0067]
According to another preferable aspect of the present invention, in the configuration of the preferable aspect 14, the pitch motion state quantity on the spring is configured to be the acceleration / deceleration of the vehicle (Preferable aspect 15).
[0068]
According to another preferable aspect of the present invention, in the configuration of the preferable aspect 15, the vehicle acceleration / deceleration is configured to be an estimated acceleration / deceleration estimated based on an acceleration / deceleration operation amount by a driver. (Preferred embodiment 16).
[0069]
According to another preferred embodiment of the present invention, in the configuration of the preferred embodiment 16, the acceleration / deceleration operation amount by the driver is configured to be a stroke of the brake pedal (preferred embodiment 17).
[0070]
According to another preferred embodiment of the present invention, in the configuration of the preferred embodiment 16, the acceleration / deceleration operation amount by the driver is the throttle opening speed (preferred embodiment 18).
[0071]
DETAILED DESCRIPTION OF THE INVENTION
The present invention will now be described in detail with reference to a few preferred embodiments with reference to the accompanying drawings.
[0072]
First embodiment
FIG. 1 is a schematic configuration diagram showing a first preferred embodiment of an attenuation coefficient control apparatus according to the present invention.
[0073]
In FIG. 1, 10FL and 10FR indicate the left and right front wheels of the vehicle 12, respectively, and 10RL and 10RR indicate the left and right rear wheels, respectively. The left and right front wheels 10FL and 10FR, which are steered wheels, are steered via tie rods 18L and 18R by a rack and pinion type power steering device 16 that is driven in response to turning of the steering wheel 14 by the driver.
[0074]
Shock absorbers 22FL to 22RR with variable damping coefficients are arranged between the unsprung wheels 10FL to 10RR and the vehicle body 20 as a sprung respectively, and the damping coefficient Ci (i = fl) of each shock absorber. , Fr, rl, rr) are controlled by the electric control device 24 when the vehicle turns as will be described later.
[0075]
The electric control device 24 includes signals indicating the strokes Xi (i = fl, fr, rl, rr) of the wheels 10FL to 10RR from the vehicle height sensors 26FL, 26FR, 26RL, 26RR, and the lateral acceleration Gy of the vehicle body from the lateral acceleration sensor 28. , A signal indicating the vehicle yaw rate γ from the yaw rate sensor 30, a signal indicating the vehicle speed V from the vehicle speed sensor 32, and a signal indicating the steering angle δ from the steering angle sensor 34.
[0076]
Although not shown in detail in the figure, the electric control device 24 has a general configuration in which, for example, a CPU, a ROM, a RAM, and an input / output port device are connected to each other by a bidirectional common bus. Includes a microcomputer. Further, the vehicle height sensors 26FL to 26RR detect the wheel stroke Xi with the bounce direction of the wheel as positive, and the lateral acceleration sensor 28 and the steering angle sensor 34 detect the lateral acceleration and the steering angle with the left turn direction of the vehicle as positive, respectively. .
[0077]
The electric control device 24 controls the damping coefficients of the shock absorbers 22FL to 22RR based on the vehicle models of the front wheels and the rear wheels shown in FIGS. 8A and 8B, respectively. In particular, the electric control device 24 of this embodiment determines whether or not the vehicle is in a transient turning state based on the lateral acceleration Gy according to the flowcharts shown in FIGS. 2 and 3 as described later, and the vehicle is in a steady turning state. When the vehicle is in a state of turning, the damping coefficient Ci of the shock absorber of each wheel is controlled to a hard damping coefficient High set in advance. The damping coefficient of the shock absorber is controlled to be higher than the damping coefficient on the outside of the turn. The damping coefficient of each shock absorber is controlled so as to increase, thereby lowering the vehicle height during transient turning and lowering the center of gravity of the vehicle body.
[0078]
The electric control device 24 calculates a reference yaw rate γt of the vehicle based on the vehicle speed V and the steering angle δ, calculates a deviation Δγ between the actual yaw rate γ and the reference yaw rate γt of the vehicle, and the vehicle is oversteered based on the deviation Δγ. Whether or not the vehicle is in an oversteer state or understeer state. When the vehicle is in an oversteer state or understeer state, a predetermined distance correction amount ΔLf, ΔLr and each virtual shock on the front wheel side and the rear wheel side so that the deviation Δγ is reduced. Absorber damping coefficient correction amounts ΔCgf, ΔCgr, ΔCf, ΔCr are calculated, and the actual damping coefficient of each shock absorber is controlled based on the calculation result.
[0079]
Next, the control of the attenuation coefficient in the illustrated first embodiment will be described with reference to the flowcharts shown in FIGS. Note that the control according to the flowcharts shown in FIGS. 2 and 3 is started by closing an ignition switch (not shown), and is repeatedly executed at predetermined time intervals.
[0080]
First, in step 10, a signal indicating the stroke Xi of each wheel is read, and in step 20, filter processing is performed to remove noise components from the detected yaw rate γ. The later yaw rate γf is calculated.
[0081]
In step 30, the actual steering angle δf of the front wheels is calculated based on the steering angle δ, the target yaw rate γe is calculated according to the following equation 30 using H as the wheel base and Kh as the stability factor, and T is the time constant. The vehicle reference yaw rate γt based on the vehicle speed V and the steering angle δ is calculated according to the following equation 31 using s as a Laplace operator. The target yaw rate γe may be calculated in consideration of the lateral acceleration Gy of the vehicle so as to take into account the dynamic yaw rate.
γe = Vδf/ (1 + KhV2) H (30)
γt = γe / (1 + Ts) (31)
[0082]
In step 40, the yaw rate deviation Δγ, that is, the deviation between the filtered yaw rate γf and the reference yaw rate γt is calculated according to the following equation 32.
Δγ = γf−γt (32)
[0083]
In step 50, it is determined whether or not the filtered yaw rate γf is a positive value, that is, whether or not the vehicle is in a left turn state. If negative determination is made, step 70 is performed. If the determination is affirmative, the process proceeds to step 60.
[0084]
In step 60, it is determined whether or not the yaw rate deviation Δγ is a positive value, that is, whether or not the vehicle is in an oversteer state. If an affirmative determination is made, the process proceeds to step 80. When the determination is made, the process proceeds to step 90.
[0085]
Similarly, in step 70, it is determined whether or not the yaw rate deviation Δγ is a negative value, that is, whether or not the vehicle is in an oversteer state. If an affirmative determination is made, the process proceeds to step 80. When a negative determination is made, the routine proceeds to step 90.
[0086]
In step 80, based on the yaw rate deviation Δγ, a predetermined distance correction amount ΔLf of the vehicle model on the front wheel side from the map corresponding to the graph shown in the first and fourth quadrants of FIG. 4, and the vehicle model on the rear wheel side. A predetermined distance correction amount ΔLr, a front wheel side virtual shock absorber 122F damping coefficient correction amount ΔCgf, a front wheel side virtual shock absorber 124F damping coefficient correction amount ΔCf, a rear wheel side virtual shock absorber 122R A damping coefficient correction amount ΔCgr and a damping coefficient correction amount ΔCr of the virtual shock absorber 124R on the rear wheel side are calculated.
[0087]
Similarly, in step 90, correction amounts ΔLf, ΔLr for predetermined distances and correction amounts ΔCgf, ΔCf for predetermined coefficients are obtained from maps corresponding to the graphs shown in the second and third quadrants of FIG. 4 based on the yaw rate deviation Δγ. , ΔCgr, ΔCr are calculated.
[0088]
In step 100, Lfo and Lro are set as the basic predetermined distances of the front and rear vehicle models, respectively, and Cgfo, Cgro, Cfo, and Cro are the virtual shock absorber 122F on the front wheel side and the rear wheel side, respectively. As basic damping coefficients set in advance for the virtual shock absorber 122R, the front wheel side virtual shock absorber 124F, and the rear wheel side virtual shock absorber 124R, predetermined distances Lf, Lr and The damping coefficient Cgf, Cgr, Cf, Cr of the virtual shock absorber is calculated.
[0089]
Lf = Lfo + ΔLf (33)
Lr = Lro + ΔLr (34)
Cgf = Cgfo + ΔCgf (35)
Cgr = Cgro + ΔCgr (36)
Cf = Cfo + ΔCf (37)
Cr = Cro + ΔCr (38)
[0090]
In step 110, for example, the deviations ΔLfa and ΔLra between the previous values and the current values of the predetermined distances Lf and Lr are calculated, and the absolute values of the deviations ΔLfa and ΔLra exceed the reference value Lo (positive constant). If the current value is corrected, the current value is corrected so that the absolute value of the deviation becomes Lo, so that the rate of change of the predetermined distances Lf and Lr is limited.
[0091]
In step 120, for example, the deviations ΔCgfa and ΔCgra between the previous values and the current values of the damping coefficients Cgf and Cgr are calculated, and the absolute values of the deviations ΔCgfa and ΔCgra exceed the reference value Cgo (positive constant). In some cases, the current value is corrected so that the absolute value of the deviation becomes Cgo, so that the rate of change of the attenuation coefficients Cgf and Cgr is limited. Also, for example, deviations ΔCfa and ΔCra between the previous values and the current values of the damping coefficients Cf and Cr are calculated, and when the absolute values of the deviations ΔCfa and ΔCra exceed the reference value Co (positive constant), the absolute value of the deviation By correcting the current value so that becomes Co, processing for limiting the rate of change of the attenuation coefficients Cf and Cr is performed, and the process proceeds to step 620.
[0092]
In step 620, it is determined whether or not the absolute value of the lateral acceleration Gy exceeds a reference value Gyo (positive constant) as a control threshold value, that is, the shock absorber damping coefficient when the wheel turns. It is determined whether or not control is necessary. When an affirmative determination is made, the process proceeds to step 640, and when a negative determination is made, the process proceeds to step 630.
[0093]
In step 630, the damping coefficient of each wheel's shock absorber is set according to a normal control routine when the vehicle is not turning, and then the routine proceeds to step 780. In this case, the control of the attenuation coefficient may be performed in any manner known in the art.
[0094]
In step 640, the time differential value ΔGy of the lateral acceleration Gy is calculated, and it is determined whether or not the absolute value of the time differential value ΔGy exceeds the reference value ΔGyo (positive constant). It is determined whether or not the vehicle is in a turning state. If an affirmative determination is made, the process proceeds to step 660. If a negative determination is made, the shock absorber damping coefficient Ci of each wheel is set in advance in step 650. After the hard attenuation coefficient “High” is set, the process proceeds to step 780.
[0095]
In step 660, the time differential value (stroke speed) Xid (i = fl, fr, rl, rr) of the stroke Xi of each wheel is calculated, and in step 670, whether the lateral acceleration Gy is positive or not. In other words, it is determined whether the vehicle is in a left turn state. If an affirmative determination is made, the process proceeds to step 680, and if a negative determination is made, the process proceeds to step 690.
[0096]
In step 680, the stroke speed Xfind of the turning inner front wheel is set to the stroke speed Xfld of the left front wheel, the stroke speed Xfoutd of the turning outer front wheel is set to the stroke speed Xfrd of the right front wheel, and the stroke speed Xrind of the turning inner rear wheel. Is set to the stroke speed Xrld of the left rear wheel, and the stroke speed Xroutd of the rear outer wheel is set to the stroke speed Xrrd of the right rear wheel.
[0097]
Similarly, at step 690, the stroke speed Xfind of the turning inner front wheel is set to the stroke speed Xfrd of the right front wheel, the stroke speed Xfoutd of the turning outer front wheel is set to the stroke speed Xfld of the left front wheel, and the stroke of the turning inner rear wheel. The speed Xrind is set to the stroke speed Xrrd of the right rear wheel, and the stroke speed Xrouted of the rear outer wheel is set to the stroke speed Xrld of the left rear wheel.
[0098]
In step 700, signGy is used as the sign of the lateral acceleration Gy, and it is determined whether or not the product of the time differential value ΔGy and signGy of the lateral acceleration is positive, that is, the magnitude of the lateral acceleration due to the turning of the vehicle increases It is determined whether or not the roll amount of the vehicle body is increasing in the process, and if an affirmative determination is made, in step 710, the turning inner front wheel, the turning outer front wheel, the turning inner rear wheel, the turning outer When the damping coefficient Cj (j = fin, fout, rin, rout) of the rear wheel shock absorber is calculated according to the equations 22 to 25, and a negative determination is made, in step 720, the damping coefficient Cj of each shock absorber is calculated. The calculation is performed according to the equations 26-29.
[0099]
In step 730, the left front wheel shock absorber damping coefficient Cfl is set to the damping inner front wheel damping coefficient Cfin, the right front wheel shock absorber damping coefficient Cfr is set to the turning outer front wheel damping coefficient Cfout, and the left rear wheel. The damping coefficient Crl of the wheel shock absorber is set to the damping coefficient Crin of the turning inner rear wheel, and the damping coefficient Crr of the right rear wheel shock absorber is set to the damping coefficient Crout of the turning outer rear wheel.
[0100]
Similarly, in step 740, it is determined whether or not the product of the lateral differential value ΔGy and signGy of the lateral acceleration is positive, and if a negative determination is made, in step 750, the turning inner front wheel, When the attenuating coefficient Cj of the shock absorber for the outer turning front wheel, the turning inner rear wheel, and the turning outer rear wheel is calculated in accordance with the above equations 22 to 25 and an affirmative determination is made, in step 760, the damping coefficient Cj of each shock absorber is calculated. The calculation is performed according to the equations 26-29.
[0101]
In step 770, the left front wheel shock absorber damping coefficient Cfl is set to the damping outer front wheel damping coefficient Cfout, the right front wheel shock absorber damping coefficient Cfr is set to the turning inner front wheel damping coefficient Cfin, and the left rear wheel. The damping coefficient Crl of the wheel shock absorber is set to the damping coefficient Crout of the rear outer wheel, and the damping coefficient Crr of the shock absorber of the right rear wheel is set to the damping coefficient Crin of the inner rear wheel.
[0102]
In step 780, the damping coefficient of each shock absorber is controlled to be the damping coefficient set in step 630, 650, 730 or 770, and then the process returns to step 10.
[0103]
Thus, according to the first embodiment shown, in step 620VehicleIn step 640, it is determined whether or not the damping coefficient of the shock absorber needs to be controlled, and in step 670, it is determined whether or not the vehicle is in a transient turning state. In step 660, 680 and 690, the stroke speed of each wheel is determined, and in steps 700 and 740, it is determined whether or not the roll amount of the vehicle body is increasing. When the roll amount of the vehicle body is in the process of increasing, the damping coefficient Cj of each shock absorber is calculated according to Equations 22-25 in steps 710 and 750, and when the roll amount of the vehicle body is in the process of decreasing, step 720 is performed. And 760, the damping coefficient Cj of each shock absorber is calculated according to equations 26-29.
[0104]
Therefore, according to the illustrated first embodiment, when the vehicle is in a transient turning state in which the roll amount of the vehicle body increases, each shock absorber is set so that the damping coefficient of the shock absorber inside the turn is higher than the damping coefficient outside the turn. When the vehicle is in a transient turning state where the roll amount of the vehicle body decreases, the damping coefficient of each shock absorber is set so that the damping coefficient of the shock absorber outside the turn is higher than the damping coefficient inside the turn. Therefore, it is possible to improve vehicle motion performance during transient turning by lowering the vehicle height and lowering the center of gravity of the vehicle body.
[0105]
Further, according to the first embodiment shown in the drawing, the damping coefficient of the left and right front wheel shock absorbers and the damping coefficient of the left and right rear wheel shock absorbers are controlled independently of each other. By appropriately setting Wf and Wr and appropriately setting a map (FIG. 4) for calculating the correction amounts ΔLf and ΔLr, ΔCgf and ΔCgr, ΔCf and ΔCr, the front and rear of the vehicle body during the transient turning of the vehicle By controlling the orientation of the direction, for example, the nose dive of the vehicle body at the beginning of turning can be reduced, or the nose lift of the vehicle body at the end of turning can be reduced.
[0106]
Further, according to the first embodiment shown in the figure, whether or not the roll amount of the vehicle body is in the increasing process or decreasing process is determined based on the lateral acceleration Gy of the vehicle body, and is detected by, for example, the vehicle height sensors 26FL to 26RR. The actual roll amount of the vehicle body is calculated based on the stroke Xi of each wheel, and the response is compared with the case where it is determined whether the vehicle body roll amount is in an increasing process or a decreasing process based on the actual roll amount. The damping coefficient of each shock absorber can be controlled with good performance.
[0107]
Since steps 620 to 780 are common in the first to fourth embodiments, the above-described effects can be obtained in the second to fourth embodiments described later.
[0108]
In particular, according to the first embodiment shown in the drawing, the reference yaw rate γt of the vehicle is calculated in step 30, the deviation Δγ between the actual yaw rate γf and the reference yaw rate γt is calculated in step 40, and steps 50 to 70 are performed. It is determined whether the vehicle is in an oversteer state or an understeer state. When the vehicle is in an oversteer state, a predetermined distance Lf of the vehicle model on the front wheel side and a virtual on the front wheel side are determined in steps 80 and 100. The shock absorber damping coefficients Cgf and Cf are corrected to increase according to the deviation Δγ, and the predetermined distance Lr of the rear wheel vehicle model and the rear shock absorber damping coefficients Cgr and Cr are the deviation Δγ. The correction is reduced according to the above.
[0109]
Conversely, when the vehicle is in the understeer state, in steps 90 and 100, the predetermined distance Lf of the vehicle model on the front wheel side and the damping coefficients Cgf and Cf of the virtual shock absorber on the front wheel side are reduced and corrected in accordance with the deviation Δγ. At the same time, the predetermined distance Lr of the vehicle model on the rear wheel side and the damping coefficients Cgr and Cr of the virtual shock absorber on the rear wheel side are increased and corrected in accordance with the deviation Δγ.
[0110]
Therefore, according to the first embodiment, when the vehicle is in the oversteer state, the roll rigidity on the front wheel side is increased and the roll rigidity on the rear wheel side is reduced according to the degree of the oversteer state, and the vehicle is in the understeer state. Sometimes, the roll rigidity on the front wheel side is reduced and the roll rigidity on the rear wheel side is increased in accordance with the degree of the understeer state, so that it is possible to reduce the steering change of the vehicle and improve the steering stability of the vehicle.
[0111]
Further, according to the first embodiment shown in the figure, the damping coefficients Cgf, Cf, Cgr, Cr of the virtual shock absorber are also corrected in accordance with the deviation Δγ in addition to the predetermined distances Lf and Lr of the vehicle model. The actual shock absorber damping coefficient can be accurately controlled as compared with the case where only a predetermined distance of the vehicle model is corrected to be increased or decreased according to the deviation Δγ.
[0112]
According to the illustrated first embodiment, the rate of change of the predetermined distances Lf and Lr is limited in step 110, and the rate of change of the damping coefficients Cgf, Cf, Cgr, Cr is limited in step 120. Therefore, it is possible to surely prevent a sudden change in the damping force of the shock absorber and a deterioration in the ride comfort of the vehicle due to this, as compared with the case where the change rate limiting process is not performed.
[0113]
In the illustrated first embodiment, the yaw rate γ of the vehicle is detected by the yaw rate sensor 30, but the wheel speeds Vwfl and Vwfr of the left and right front wheels, which are the steering wheels, are detected, and Tr is The vehicle tread may be calculated according to the following equation 39 based on the wheel speed.
γ = (Vwfr−Vwfl) / Tr (39)
[0114]
In the illustrated first embodiment, the damping coefficients Cgf, Cf, Cgr, Cr of the virtual shock absorber are corrected in accordance with the deviation Δγ in addition to the predetermined distances Lf and Lr of the vehicle model. However, the increase / decrease correction of the attenuation coefficients Cgf, Cf, Cgr, Cr may be omitted. In the illustrated embodiment, the increase / decrease correction amounts of the attenuation coefficients Cgf and Cgr are set larger than the increase / decrease correction amounts of the attenuation coefficients Cf and Cr when the same deviation Δγ is observed. The increase / decrease correction amount of Cr may be set larger than the increase / decrease correction amounts of the attenuation coefficients Cgf and Cgr, and further, the increase / decrease correction of one of the attenuation coefficients Cgf, Cgr and the attenuation coefficients Cf, Cr may be omitted.
[0115]
Further, in the first embodiment shown in the figure, the target yaw rate γe of the vehicle is calculated according to the above equation 30, but corresponds to the graph shown in FIG. 5 based on the vehicle speed V and the steering angle δ. May be calculated from the map.
[0116]
Second embodiment
FIG. 9 is a flowchart showing the first half of the damping coefficient control routine in the second preferred embodiment of the damping coefficient control apparatus according to the present invention.
[0117]
Although not shown in the figure, the electric control device 24 of the second embodiment is not input with a signal indicating the yaw rate γ of the vehicle, a signal indicating the vehicle speed V, and a signal indicating the steering angle δ. A signal indicating the throttle opening degree Th of the engine and a signal indicating the depression stroke Sb of the brake pedal are also input from the brake sensor.
[0118]
In step 210 of the damping coefficient control routine of the second embodiment, a signal indicating the stroke Xi of each wheel is read, and in step 220, for example, as a time differential value of the throttle opening Th. The throttle opening speed Vt is calculated.
[0119]
In step 230, based on the throttle opening speed Vt and the vehicle speed V, the map corresponding to the graph shown in FIG. 10 (front wheel drive vehicle) or FIG. 11 (rear wheel drive vehicle) is used for a predetermined distance Lf on the front wheel side. The distribution ratio Ka is calculated.
[0120]
In step 240, the distribution ratio Kb to the predetermined distance Lf on the front wheel side is calculated from the map corresponding to the graph shown in FIG. 12 based on the brake stroke Sb.
[0121]
In step 250, a predetermined distance Lf on the front wheel side and a distance Lr on the rear wheel side are calculated according to the following equations 40 and 41, respectively, and then the process proceeds to step 620.
Lf = KaKbLfo (40)
Lr = (1-Ka) (1-Kb) Lro (41)
[0122]
Thus, according to the second embodiment shown in the figure, the throttle opening speed Vt is calculated in step 220, and the distribution to the predetermined distance Lf on the front wheel side based on the throttle opening speed Vt and the vehicle speed V in step 230. The ratio Ka is calculated, and in step 250, a predetermined distance Lf on the front wheel side and a predetermined distance Lr on the rear wheel side are calculated based on the distribution ratio of the front and rear wheels based on the distribution ratio Ka.
[0123]
In general, in the case of a front-wheel drive vehicle, the steering characteristic of the vehicle changes to the understeer side due to the driving force of the front wheel during acceleration of the vehicle, and conversely, in the case of a rear-wheel drive vehicle, it results from the driving force of the rear wheel. However, according to the illustrated second embodiment, when the vehicle is a front-wheel drive vehicle, the rear-wheel side steer characteristic with respect to a predetermined distance Lf on the front-wheel side is accelerated when the vehicle is accelerated. Since the ratio of the predetermined distance Lr is increased, the change in the steering characteristic toward the understeer side of the vehicle is reduced, and when the vehicle is a rear-wheel drive vehicle, the predetermined distance Lr on the rear wheel side when the vehicle is accelerated. Since the ratio of the predetermined distance Lf on the front wheel side to the vehicle is increased, the change in the steering characteristic toward the oversteer side of the vehicle is reduced, and thus the steering stability of the vehicle can be improved.
[0124]
Further, according to the second embodiment shown in the drawing, the distribution ratio Kb to the predetermined distance Lf on the front wheel side is calculated based on the brake stroke Sb in step 240, and the front and rear wheels based on the distribution ratio Kb are calculated in step 250. Based on the distribution ratio, a predetermined distance Lf on the front wheel side and a predetermined distance Lr on the rear wheel side are calculated.
[0125]
Therefore, according to the second embodiment, it is possible to reduce the change in the steering characteristic toward the oversteer side caused by the load movement forward of the vehicle during braking of the vehicle, thereby improving the steering stability of the vehicle. .
[0126]
Further, according to the second embodiment shown in the figure, the acceleration / deceleration of the vehicle is estimated based on the brake stroke Sb that is the amount of braking operation of the driver and the throttle opening speed that is the amount of acceleration operation of the driver. The damping coefficient of the shock absorber of each wheel can be controlled with higher responsiveness than when acceleration / deceleration is detected by, for example, a longitudinal acceleration sensor.
[0127]
In the second embodiment shown in the figure, the vehicle acceleration as the vehicle pitching state quantity is estimated based on the throttle opening speed Vt. However, the vehicle acceleration is, for example, that of an automatic transmission. You may estimate based on the output torque etc. of a torque converter. The vehicle deceleration as the vehicle pitching state quantity is estimated on the basis of the brake stroke Sb. The vehicle deceleration is not, for example, shown in the drawing of the brake pedal or in the brake master cylinder. May be estimated based on the pressure.
[0128]
Third embodiment
FIG. 13 is a flowchart showing the first half of the damping coefficient control routine in the third preferred embodiment of the damping coefficient control apparatus according to the present invention.
[0129]
In step 260 of the damping coefficient control routine of the third embodiment, a signal indicating the stroke Xi of each wheel is read, and in step 270, for example, steering is performed as a time differential value of the steering angle δ. The angular velocity δd is calculated, and the lateral jerk magnitude Jy of the vehicle is calculated according to the following equation 42, with K1 and K2 being positive constants.
Jy = | K1δV²+ K2δV²｜ …… (42)
[0130]
In step 280, a predetermined distance Lf on the front wheel side and a predetermined distance Lr on the rear wheel side are calculated from the map corresponding to the graph shown in FIG. 14 based on the magnitude Jy of the lateral jerk of the vehicle. Proceed to step 620.
[0131]
Thus, according to the illustrated third embodiment, since the predetermined distances Lf and Lr are variably set so as to increase as the lateral jerk Jy of the vehicle increases, the vehicle is more likely to change the roll angle of the vehicle body. Thus, the change in the posture of the roll of the vehicle body during the transient turning of the vehicle can be appropriately suppressed according to the turning situation of the vehicle.
[0132]
Further, according to the third embodiment shown in the figure, when the lateral jerk magnitude Jy of the vehicle is extremely large, the predetermined distances Lf and Lr are gradually reduced as the lateral jerk magnitude Jy increases. Therefore, when a very sudden transient turning operation is performed, the vehicle body rolls, and accordingly, the driver can be alerted to take appropriate measures such as decelerating.
[0133]
Fourth embodiment
FIG. 15 is a flowchart showing the first half of a damping coefficient control routine in the fourth preferred embodiment of the damping coefficient control apparatus according to the present invention.
[0134]
Although not shown in the figure, a signal indicating the vehicle yaw rate γ, a signal indicating the vehicle speed V, and a signal indicating the steering angle δ are not input to the electric control device 24 of the fourth embodiment. Signals indicating the pressure Pui (i = fl, fr, rl, rr) in the cylinder upper chamber and the pressure Pli (i = fl, fr, rl, rr) in the cylinder lower chamber of the shock absorbers 22FL-22RR detected by the pressure sensor Is also entered.
[0135]
In step 310 of the damping coefficient control routine of the fourth embodiment, a signal indicating the stroke Xi of each wheel is read, and in step 320, the shocks are respectively expressed according to the following equations 43 to 46. A differential pressure Pdfi (i = fl, fr, rl, rr) between the pressure Pui in the cylinder upper chamber and the pressure Pli in the cylinder lower chamber of the absorbers 22FL to 22RR is calculated.
[0136]
Pdfl = Pufl−Plfl (43)
Pdfr = Pufr−Plfr (44)
Pdfrl = Purl-Plrl (45)
Pdfrr = Purr−Plrr (46)
[0137]
In step 330, the maximum differential value Pdf on the front wheel side is calculated according to the following equation 47, with MAX being the larger value in (), and in step 340, based on the representative differential pressure Pdff on the front wheel side. A predetermined distance Lf on the front wheel side is calculated from the map corresponding to the graph shown in FIG.
Pdff = MAX (Pdfl, Pdfr) (47)
[0138]
In step 350, the representative differential pressure Pdfr on the rear wheel side is calculated according to the following equation 48. In step 360, the map corresponding to the graph shown in FIG. 17 is based on the representative differential pressure Pdfr on the rear wheel side. Further, a predetermined distance Lr on the rear wheel side is calculated, and then the routine proceeds to step 620.
Pdfr = MAX (Pdfrl, Pdfrr) (48)
[0139]
Thus, according to the fourth embodiment, in step 320, the differential pressure Pdfi between the pressure Pui in the cylinder upper chamber and the pressure Pli in the cylinder lower chamber of each of the shock absorbers 22FL to 22RR is calculated. The larger value of the differential pressures Pdfl and Pdfr is set as the representative differential pressure Pdff on the front wheel side. In step 340, a predetermined value on the front wheel side is set so as to increase as the representative differential pressure increases based on the representative differential pressure Pdff on the front wheel side. The distance Lf is calculated. In step 350, the larger value of the differential pressures Pdfrl and Pdfrr is set as the representative differential pressure Pdfr on the rear wheel side. In step 360, the larger the representative differential pressure is based on the representative differential pressure Pdfr on the rear wheel side. The predetermined distance Lr on the rear wheel side is calculated so that
[0140]
Therefore, according to the fourth embodiment, the greater the representative differential pressure, in other words, the higher the relative displacement speed between the vehicle body and the wheels, the greater the predetermined distance of the vehicle model is increased, thereby increasing the actual shock absorber. As the damping coefficient of the vehicle is increased, it is possible to reduce changes in the posture of the vehicle body caused by disturbance during transient turning of the vehicle, and to improve the ground performance of the vehicle and improve the motion performance during transient turning of the vehicle Can be made.
[0141]
Fifth embodiment
FIG. 19 is a flowchart showing the first half of the damping coefficient control routine in the fifth preferred embodiment of the damping coefficient control apparatus according to the present invention.
[0142]
In step 410 of the damping coefficient control routine of the fifth embodiment, a signal indicating the stroke Xi of each wheel is read, and in step 420, the product of the lateral acceleration Gy, the vehicle speed V, and the yaw rate γ. The deviation of the lateral acceleration, that is, the lateral slip acceleration Vyd of the vehicle is calculated as the deviation Gy−Vγ from Vγ, and the lateral slip speed Vy of the vehicle body is calculated by integrating the lateral slip acceleration Vyd. The slip angle β of the vehicle body is calculated as the ratio Vy / Vx of the vehicle body side slip velocity Vy to the above.
[0143]
In step 430, for example, the rate of change βd of the slip angle of the vehicle body is calculated as a time differential value of the slip angle β of the vehicle body, and the graph shown in FIG. 20 based on the slip angle β of the vehicle body and the rate of change βd thereof. In step 440, the correction coefficient Kf for the predetermined distance Lf on the front wheel side and the correction coefficient Kr for the predetermined distance Lr on the rear wheel side are determined. Set to 1.
[0144]
In the graph of FIG. 20, the area A shows the normal stable running state of the vehicle, the area B shows the area where the vehicle is in a spin tendency state, and the area C shows a state where the vehicle is in a drift-out tendency state. A certain region is shown, and a region D shows a region where the turning behavior of the vehicle diverges.
[0145]
In step 450, it is determined whether or not the vehicle is in the region B. If a negative determination is made, the process proceeds to step 470, and if an affirmative determination is made, the slip angle β of the vehicle body is determined in step 460. The correction coefficient Kf is calculated from the map corresponding to the graph shown in FIG.
[0146]
In step 470, it is determined whether or not the vehicle is in the region C. If a negative determination is made, the process proceeds to step 490. If an affirmative determination is made, the vehicle slips in step 480. Based on the absolute value of the angle β, the correction coefficient Kr is calculated from the map corresponding to the graph shown in FIG.
[0147]
In step 490, it is determined whether or not the vehicle is in the region D. If a negative determination is made, the process proceeds to step 510. If an affirmative determination is made, the slip angle β of the vehicle body is determined in step 500. Based on the absolute value ofFrom the map corresponding to the graph shown in FIG.Correction coefficients Kf and Kr are calculated.
[0148]
In step 510, a predetermined distance Lf on the front wheel side and a predetermined distance Lr on the rear wheel side are calculated according to the following equations 49 and 50, respectively, and in step 520, step 110 in the first embodiment is performed. In the same manner as in the case of the above, the limiting process of the rate of change of the predetermined distances Lf and Lr is performed, and then the process proceeds to step 620.
Lf = KfLfo (49)
Lr = KrLro (50)
[0149]
Thus, according to the fifth embodiment shown in the figure, the slip angle β of the vehicle body is calculated in step 420, the rate of change βd of the slip angle of the vehicle body is calculated in step 430, and the slip angle β of the vehicle body is calculated. When the vehicle is in the region B, that is, in the state of spin tendency, in step 450, it is determined in which region the vehicle is in the turning behavior based on the rate of change βd. The predetermined distance Lf of the vehicle model of the front wheels is increased so as to increase as the absolute value of the slip angle β of the vehicle body increases, thereby increasing the roll rigidity of the vehicle on the front wheel side.
[0150]
If it is determined in step 470 that the vehicle is in the region C, that is, in a drift-out tendency state, in step 480, the vehicle model of the rear wheel is increased so that the absolute value of the slip angle β of the vehicle body increases. The predetermined distance Lr is increased, thereby increasing the roll rigidity of the vehicle on the rear wheel side.
[0151]
Further, if it is determined in step 490 that the vehicle is in the region D, that is, the region in which the turning behavior of the vehicle diverges, in step 500, the front wheel and the rear are increased so that the absolute value of the slip angle β of the vehicle body increases. The predetermined distances Lf and Lr of the wheeled vehicle model are increased, thereby increasing the roll stiffness of the entire vehicle.
[0152]
Therefore, according to the fifth embodiment, the turning behavior of the vehicle at the time of the transient turning of the vehicle is determined, and the damping coefficient of the actual shock absorber is appropriately controlled according to the determination result. Deterioration of turning behavior such as drift-out can be effectively suppressed.
[0153]
Further, according to the fifth embodiment, since the rate of change of the predetermined distances Lf and Lr is limited in step 520, the damping force of the shock absorber is compared with a case where the rate of change limiting process is not performed. It is possible to reliably prevent a sudden change in the vehicle speed and a deterioration in the ride comfort of the vehicle due to this.
[0154]
Although the present invention has been described in detail with respect to specific embodiments, the present invention is not limited to the above-described embodiments, and various other embodiments are possible within the scope of the present invention. It will be apparent to those skilled in the art.
[0155]
For example, in each of the above-described embodiments, the damping coefficient of the shock absorber inside the turning is controlled to be relatively higher than the damping coefficient of the shock absorber outside the turning in the process of increasing the body roll amount. In the process of decreasing the roll amount, the damping coefficient of the shock absorber on the outside of the turn is controlled to be higher than the damping coefficient of the shock absorber on the inside of the turn. In this case, the possibility that the behavior of the vehicle becomes unstable is lower than the process of increasing the roll amount of the vehicle body (in the initial stage of turning), so the damping coefficient of the shock absorber outside the turn is turned in the process of decreasing the roll amount of the vehicle body. Control to make it higher than the damping coefficient of the inner shock absorber may be omitted.
[0156]
Specifically, the calculation of the damping coefficient Cj in steps 720 and 760 is omitted, and instead, the damping coefficient Ci of each shock absorber is set to the hard damping coefficient High as in the case of step 650, for example. It may be modified to proceed to 780.
[0157]
In each of the above embodiments, the vehicle turning direction is determined based on the sign of the lateral acceleration Gy of the vehicle body. For example, Kh is a stability factor and Rg is a steering gear ratio. Based on the steering angle δ detected by the steering angle sensor 28 shown in FIG. 1 and the vehicle speed V detected by the vehicle speed sensor not shown in FIG. The lateral acceleration Gys of the vehicle may be estimated based on the estimated lateral acceleration, and may be performed based on the estimated lateral acceleration.
Gys = V²δ / [(1 + Kh V²) Rg H] (51)
[0158]
Similarly, the yaw rate γh of the vehicle may be estimated based on the vehicle speed and the steering angle, and the turning direction of the vehicle may be determined based on the sign. Thus, when the turning direction of the vehicle is determined based on the estimated lateral acceleration Gys or the estimated yaw rate γh of the vehicle body, for example, the turning direction of the vehicle determined as in the case of counter steer is opposite to the actual turning direction of the vehicle. (1) Judgment of the turning direction based on the estimated lateral acceleration Gys or estimated yaw rate γh of the vehicle body, and (2) the estimation based on the actual lateral acceleration Gy of the vehicle body or the wheel speeds of the left and right steering wheels. Turning direction based on the difference between the yaw rate γh or the internal pressure of the left and right shock absorbers, the difference between the stroke speeds of the left and right wheels, or the difference between the left and right sprung speeds. It may be modified to be adopted.
[0159]
In each of the above-described embodiments, it is determined whether the vehicle body roll amount is in the increasing process or the decreasing process based on the sign of the time differential value ΔGy of the lateral acceleration Gy of the vehicle body. For example, this determination may be made based on the estimated lateral acceleration Gys calculated according to the above equation 51, and the sign of the vehicle body roll rate calculated based on the stroke Xi detected by the vehicle height sensors 26FL to 26RR. May be made based on. In this case, the roll rate may be detected by a roll rate sensor not shown in FIG.
[0160]
In each of the above-described embodiments, the stroke speed Xid of each wheel is calculated based on the detection results of the vehicle height sensors 26FL to 26RR, but the stroke speed of each wheel is provided in the vehicle body. It may be estimated by an observer based on the vertical acceleration Gbi of the vehicle body detected by a vertical acceleration sensor (not shown), and the vehicle height sensor may be omitted.
[0161]
In the second to fifth embodiments described above, only the predetermined distances Lf and Lr are variably set. However, as in the case of the first embodiment, these embodiments are used. However, the damping coefficients Cgf, Cgr, Cf, Cr of the virtual shock absorber may be variably set according to the state quantity of the vehicle together with the predetermined distances Lf, Lr.
[0162]
In the first embodiment described above, the predetermined distances Lf, Lr and the rate of change of the damping coefficient Cgf, Cgr, Cf, Cr of the virtual shock absorber are limited. In the fifth embodiment, Although the change rates of the predetermined distances Lf and Lr are limited, these change rate limiting processes may be omitted, and the predetermined distance Lf is also used in the second to fourth embodiments. , Lr may be modified to limit the rate of change.
[0163]
Furthermore, in the second embodiment described above, the predetermined distances Lf and Lr and the ratio thereof are changed based on both the throttle opening speed Vt and the brake stroke Sb. Lf, Lr and their ratio may be modified so as to be changed based on only one of the throttle opening speed Vt and the brake stroke Sb.
[0164]
【The invention's effect】
As is apparent from the above description, according to the configuration of claim 1 of the present invention, during a transient turning of the vehicle.The first virtual shock absorber suppresses the roll displacement on the spring,Swivel on a spring by a second virtual shock absorberInsideSince the wheel side lift is suppressed,Reduces roll displacement on the springLower the center of gravity on the spring,TransientIn addition to improving the motion performance during turning, the predetermined distance of the vehicle model is variably set according to the state quantity of the vehicle. As a result, the vehicle steer characteristics change and the spring posture change during the vehicle's transient turning, compared with the case where the predetermined distance of the vehicle model is constant regardless of the state quantity of the vehicle. Suppression can be appropriately performed according to the running state of the vehicle.
[0165]
According to the second aspect of the present invention, since the vehicle model is composed of a front wheel vehicle model and a rear wheel vehicle model, the actual shock absorber attenuation coefficient on the front wheel side and the rear wheel side is determined according to the state quantity of the vehicle. Appropriately controlled, so that the change in the steering characteristics of the vehicle and the change in the posture on the spring during the transient turning of the vehicle are appropriately controlled according to the running state of the vehicle, compared to the case where there is only one vehicle model. Can be controlled.
[0166]
Further, according to the configuration of claim 3, since the predetermined distance of the vehicle model is variably set according to the turning behavior of the vehicle, the actual shock absorber attenuation coefficient is appropriately controlled according to the turning behavior of the vehicle, This makes it possible to appropriately suppress the change in the steering characteristic of the vehicle during the transient turning of the vehicle, as compared with the case where the predetermined distance of the vehicle model is constant regardless of the turning behavior of the vehicle.
[0167]
According to the configuration of claim 4, since the predetermined distance of the vehicle model is variably set according to the lateral jerk of the vehicle, the damping coefficient of the actual shock absorber is appropriately controlled according to the lateral jerk of the vehicle, As a result, the roll posture change on the spring during the transient turning of the vehicle can be appropriately suppressed as compared with the case where the predetermined distance of the vehicle model is constant regardless of the lateral jerk of the vehicle.
[0168]
According to the fifth aspect of the present invention, since the predetermined distance of the vehicle model is variably set in accordance with the differential pressure between the two cylinder chambers of the actual shock absorber, the damping coefficient of the actual shock absorber is set to the two cylinders. Appropriate control is performed according to the differential pressure of the indoor pressure, and thereby the posture change on the spring can be appropriately controlled.
[0169]
According to the configuration of claim 6, since the ratio of the predetermined distance between the two vehicle models is variably set according to the turning behavior of the vehicle, the actual damping coefficient of the shock absorber is appropriately set according to the turning behavior of the vehicle. Therefore, regardless of the turning behavior of the vehicle, the change of the pitch posture on the spring and the change in the pitch posture on the spring during the transient turning of the vehicle are compared with the case where the ratio of the predetermined distance between the two vehicle models is constant. It is possible to appropriately suppress the change in the steering characteristic due to the above.
[Brief description of the drawings]
FIG. 1 is a schematic configuration diagram showing a first preferred embodiment of an attenuation coefficient control apparatus according to the present invention.
FIG. 2 is a flowchart showing the first half of an attenuation coefficient control routine in the first embodiment.
FIG. 3 is a flowchart showing the second half of an attenuation coefficient control routine in the first embodiment.
FIG. 4 is a graph showing a relationship between a yaw rate deviation Δγ, correction amounts ΔLf and ΔLr for a predetermined distance, and attenuation coefficient correction amounts ΔCgf, ΔCgr, ΔCf, and ΔCr.
FIG. 5 is a graph showing a relationship between a steering angle δ and a vehicle speed V and a target yaw rate γe of the vehicle.
FIG. 6 is an explanatory diagram showing a two-wheel model of an actual vehicle.
FIG. 7 is an explanatory diagram showing a virtual model in which a virtual shock absorber is disposed inside the turning of the vehicle.
FIG. 8 is an explanatory diagram showing a virtual model on the front wheel side and the rear wheel side in which virtual shock absorbers are disposed inside the turning of the vehicle.
FIG. 9 is a flowchart showing the first half of an attenuation coefficient control routine in the second embodiment.
FIG. 10 is a graph showing a relationship between a throttle opening speed Vt and a distribution ratio Ka with respect to a predetermined distance Lf on the front wheel side for a front wheel drive vehicle.
FIG. 11 is a graph showing a relationship between a throttle opening speed Vt and a distribution ratio Ka with respect to a predetermined distance Lf on the front wheel side for a rear wheel drive vehicle.
FIG. 12 is a graph showing a relationship between a brake stroke Sb and a distribution ratio Kb with respect to a predetermined distance Lf on the front wheel side for a normal vehicle.
FIG. 13 is a flowchart showing the first half of an attenuation coefficient control routine in the third embodiment.
FIG. 14 is a graph showing a relationship between a lateral jerk magnitude Jy of a vehicle and predetermined distances Lf and Lr.
FIG. 15 is a flowchart showing the first half of an attenuation coefficient control routine in the fourth embodiment.
FIG. 16 is a graph showing a relationship between a representative differential pressure Pdff on the front wheel side and a predetermined distance Lf on the front wheel side.
FIG. 17 is a graph showing a relationship between a representative differential pressure Pdfr on the rear wheel side and a predetermined distance Lr on the rear wheel side.
FIG. 18 is a graph showing the relationship between the stroke speed Xid of each wheel and the differential pressure Pdfi.
FIG. 19 is a flowchart showing the first half of an attenuation coefficient control routine in the fifth embodiment.
FIG. 20 is a graph showing the relationship between the slip angle β of a vehicle and its differential value βd and each region.
FIG. 21 is a graph showing a relationship between an absolute value of a vehicle slip angle β and a correction coefficient Kf with respect to a predetermined distance Lf on the front wheel side.
FIG. 22 is a graph showing the relationship between the absolute value of the vehicle slip angle β and the correction coefficient Kr with respect to a predetermined distance Lr on the rear wheel side.
FIG. 23 is a graph showing the relationship between the absolute value of the slip angle β of the vehicle and the correction coefficients Kf and Kr.
[Explanation of symbols]
14 ... Steering wheel
16 ... Power steering device
20 ... Body
24. Electric control device
26FL-26RR ... Vehicle height sensor
28 ... Lateral acceleration sensor
30 ... Yaw rate sensor
32 ... Vehicle speed sensor
34 ... Steering angle sensor
110 ... Body
112L, 112R ... wheels
114L, 114R ... suspension spring
116L, 116R ... Shock absorber
122, 124 ... Shock absorber

Claims (6)

A vehicle damping coefficient control device provided with an actual shock absorber with variable damping coefficient corresponding to each wheel, means for detecting the turning state of the vehicle, means for detecting the state quantity of the vehicle, and the vehicle with the virtual position spaced a predetermined distance the vehicle transverse direction than on the spring to the center of gravity of the sprung based on the turning state to the side that is estimated to be lifted with a virtual pivot center on the spring, the A first virtual shock absorber that generates a damping force that suppresses roll displacement on the spring around a virtual swing center, and a virtual wheel positioned below the virtual swing center and the spring a vehicle model having a second virtual shock absorber that generates a vertical damping force in between, the distance setting means for variably setting the predetermined distance in accordance with the state quantity of the vehicle, at least the provisionally Means for calculating a target damping coefficient of the actual shock absorber based on the damping coefficient and the predetermined distance; and means for controlling the damping coefficient of the actual shock absorber based on the target damping coefficient. Vehicle damping coefficient control device. 2. The vehicle damping coefficient control device according to claim 1, wherein the vehicle model includes a front-wheel vehicle model and a rear-wheel vehicle model. The means for detecting the state quantity of the vehicle detects the turning behavior of the vehicle, and the distance setting means variably sets the predetermined distance according to the turning behavior of the vehicle. The vehicle damping coefficient control device described in 1. The means for detecting a state quantity of the vehicle detects lateral jerk of the vehicle, and the distance setting means variably sets the predetermined distance according to the lateral jerk of the vehicle. The vehicle damping coefficient control device described in 1. The means for detecting the state quantity of the vehicle detects a differential pressure between two cylinder chamber pressures of the actual shock absorber, and the distance setting means variably sets the predetermined distance according to the differential pressure. The vehicle damping coefficient control device according to claim 1 or 2. The means for detecting the state quantity of the vehicle detects the turning behavior of the vehicle, and the distance setting means determines the predetermined vehicle between the front wheel vehicle model and the rear wheel vehicle model according to the turning behavior of the vehicle. 3. The vehicle damping coefficient control device according to claim 2, wherein the ratio of the distances is variably set.

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