JP3847487B2

JP3847487B2 - Hydraulic power steering device

Info

Publication number: JP3847487B2
Application number: JP13895299A
Authority: JP
Inventors: 晃久梅谷; 修宮崎
Original assignee: JTEKT Corp
Current assignee: JTEKT Corp
Priority date: 1999-05-19
Filing date: 1999-05-19
Publication date: 2006-11-22
Anticipated expiration: 2019-05-19
Also published as: JP2000326859A

Description

【０００１】
【発明の属する技術分野】
本発明は、操舵補助力発生用油圧アクチュエータに作用する油圧をロータリータイプ制御弁によって制御するラックピニオン式油圧パワーステアリング装置に関する。
【０００２】
【従来の技術】
油圧パワーステアリング装置においては、油圧アクチュエータに作用する油圧を制御弁により制御することで操舵補助力を発生している。その制御弁として、筒状の第１バルブ部材と、この第１バルブ部材に操舵抵抗に応じて相対回転可能に挿入される第２バルブ部材とを有するロータリータイプのものが用いられている。
【０００３】
そのロータリータイプの制御弁においては、第１バルブ部材の内周に複数の軸方向に沿う溝が互いに周方向の間隔をおいて形成され、その第２バルブ部材の外周に複数の軸方向に沿う溝が互いに周方向の間隔をおいて形成されている。その第１バルブ部材の溝の軸方向に沿う縁と、第２バルブ部材の溝の軸方向に沿う縁との間が、両バルブ部材の相対回転角度に応じて開度が変化する複数の絞り部とされている。その制御弁を介して前記アクチュエータがポンプとタンクに接続されることで、各絞り部の開度変化に応じた操舵補助力を付与できる。
【０００４】
また、車両の高速走行時における安定性と低速走行時における旋回性を向上するため、高速走行時において操舵補助力を低減し、低速走行時において操舵補助力を増大している。そのため、運転条件に応じて操舵補助力を変更する構成が採用されている。すなわち、各絞り部は第１の組と第２の組とに組分けされ、第２の組に属する絞り部の閉鎖角度は第１の組に属する絞り部の閉鎖角度よりも大きくされ、第２の組に属する絞り部とタンクとの間の油路に、運転条件に応じて開度が変化する可変絞り部が設けられている。その可変絞り部の開度は、高速走行時に大きくされ、低速走行時に小さくされる。これにより、操舵補助力を発生させる油圧は、高速走行時においては第１の組と第２の組の双方の絞り部の開度に応じて変化し、速度が低下するに伴って第２の組の絞り部の開度の影響が小さくなる。これにより、高速走行時においては、操舵抵抗に応じた操舵トルク変化に対する油圧変化の少ない領域が大きくなり、低速走行時においては、その操舵トルク変化に対する油圧変化の少ない領域が小さくなる。
【０００５】
上記各絞り部の閉鎖角度の設定のため、その第２バルブ部材における各溝の軸方向に沿う縁は面取り部とされ、この面取り部の周方向幅は、各絞り部を全閉するのに要する両バルブ部材の相対回転角度である閉鎖角度に応じて定められている。
【０００６】
【発明が解決しようとする課題】
上記可変絞り部が開かれている中速走行状態では、第１の組に属する絞り部が閉鎖された後に、第２の組に属する絞り部の開度が可変絞り部の開度よりも小さくなるまでの間においては、その可変絞り部の開度は一定になるため、油圧アクチュエータに作用する油圧を操舵抵抗に応じて制御できない。すなわち、その間においては操舵トルク変化に対して操舵補助力発生用の油圧が追従しなくなるため、操舵フィーリングが低下する。
【０００７】
そこで、第１の組に属する絞り部の閉鎖タイミングをできるだけ遅らせることで、第１の組に属する絞り部が閉鎖された後に、第２の組に属する絞り部の開度が可変絞り部の開度よりも小さくなるまでの間を小さくすることが考えられる。
【０００８】
しかし、従来の上記各面取り部は単一の平坦面により構成されていた。そのため、第１の組に属する絞り部の閉鎖タイミングを遅らせるには、第１の組に属する絞り部の面取り部の周方向幅を大きくする必要がある。この場合、絞り機能を確保するため、第２バルブ部材の外接円の径方向に対して、第１の組に属する絞り部の面取り部を構成する平坦面に直角な方向の傾き角を小さくする必要がある。そうすると、第１の組に属する絞り部において閉鎖開始初期の流路面積が急激に小さくなる。そのため、低中速走行時の操舵開始初期に操舵補助力発生用油圧が急上昇し、操舵トルクに対する操舵補助力発生用油圧の変化割合が急変するために操舵フィーリングが低下する。また、中速走行時に操舵トルクが小さくても操舵補助力発生用油圧が大きくなることで走行安定性が低下する。さらに、高速走行時に操舵補助力発生用油圧を十分に大きくするのに必要な操舵トルクが過大になり、必要な操舵補助を行うことができないという問題がある。
【０００９】
本発明は、上記問題を解決することのできる油圧パワーステアリング装置を提供することを目的とする。
【００１０】
【課題を解決するための手段】
本発明は、操舵補助力発生用油圧アクチュエータと、そのアクチュエータに作用する油圧の制御弁とを備え、その制御弁は、筒状の第１バルブ部材と、この第１バルブ部材に操舵抵抗に応じて相対回転可能に挿入される第２バルブ部材とを有し、その第１バルブ部材の内周に複数の軸方向に沿う溝が互いに周方向の間隔をおいて形成され、その第２バルブ部材の外周に複数の軸方向に沿う溝が互いに周方向の間隔をおいて形成され、その第１バルブ部材の溝の軸方向に沿う縁と、第２バルブ部材の溝の軸方向に沿う縁との間が、両バルブ部材の相対回転角度に応じて開度が変化する複数の絞り部とされ、各絞り部の開度変化に応じた操舵補助力を付与できるように、その制御弁を介して前記アクチュエータがポンプとタンクに接続され、前記第２バルブ部材における各溝の軸方向に沿う縁は面取り部とされ、この面取り部の周方向幅は、各絞り部を全閉するのに要する両バルブ部材の相対回転角度である閉鎖角度に応じて定められ、各絞り部は第１の組と第２の組とに組分けされ、第２の組に属する絞り部の閉鎖角度は第１の組に属する絞り部の閉鎖角度よりも大きくされ、第２の組に属する絞り部とタンクとの間の油路に、運転条件に応じて開度が変化する可変絞り部が設けられている油圧パワーステアリング装置に適用される。
本発明の油圧パワーステアリング装置において、その第１の組に属する絞り部における前記面取り部は、周方向に並列する複数の平坦面から構成され、その第２バルブ部材の外接円の径方向に対して、前記各平坦面に直角な方向の傾き角は互いに相異するものとされ、前記第２バルブ部材の溝に近い方の平坦面に直角な方向は、その溝から遠い方の平坦面に直角な方向よりも、前記傾き角が大きくされていることを特徴とする。
本発明の構成によれば、第１の組に属する絞り部における面取り部を構成する複数の平坦面の中で、第２バルブ部材の溝から遠い方の平坦面に直角な方向の前記外接円の径方向に対する傾き角を小さくすることで、その第１の組に属する絞り部における絞り機能を確保しつつ上記面取り部の周方向幅を大きくできる。これにより、第１の組に属する絞り部の流路面積が最小になった後に、第２の組に属する絞り部の開度が可変絞り部の開度よりも小さくなるまでの間を小さくできる。よって、第１の組に属する絞り部の閉鎖タイミングを遅らせ、中速走行時において操舵補助力を操舵抵抗に応じ制御できない領域を小さくし、操舵トルク変化に対する操舵補助力発生用の油圧の追従性を向上できる。
しかも、第１の組に属する絞り部における面取り部を構成する複数の平坦面の中で、第２バルブ部材の溝に近い方の平坦面に直角な方向の前記外接円の径方向に対する傾き角を大きくすることで、第１の組に属する絞り部において閉鎖開始初期の流路面積が急激に小さくなるのを防止できる。これにより、低中速走行時の操舵開始初期に操舵補助力発生用油圧が急上昇することはなく、操舵トルクに対して略一定割合で操舵補助力発生用油圧を増加させて操舵フィーリングを向上できる。また、中速走行時において、操舵補助力発生用油圧が過大になるのを防止して走行安定性を向上できる。さらに、高速走行時に操舵補助力発生用油圧を十分に大きくするのに必要な操舵トルクが過大になることはなく、必要な操舵補助を行うことができる。
【００１１】
【発明の実施の形態】
図１に示す本発明の実施形態のラックピニオン式油圧パワーステアリング装置１は、車両のハンドル（図示省略）に連結される入力軸２と、この入力軸２にトーションバー６を介し連結される出力軸３を備えている。そのトーションバー６は、ピン４により入力軸２に連結され、セレーション５により出力軸３に連結されている。その入力軸２は、ベアリング８を介しバルブハウジング７により支持され、また、ブッシュ１２を介し出力軸３により支持されている。その出力軸３はベアリング１０、１１を介しラックハウジング９により支持されている。その出力軸３にピニオン１５が形成され、このピニオン１５に噛み合うラック１６に操舵用車輪（図示省略）が連結される。これにより、操舵による入力軸２の回転は、トーションバー６を介してピニオン１５に伝達され、このピニオン１５の回転によりラック１６は車両幅方向に移動し、このラック１６の移動により車両の操舵がなされる。なお、入出力軸２、３とハウジング７との間にはオイルシール４２、４３が介在する。また、ラック１６を支持するサポートヨーク４０がバネ４１の弾力によりラック１６に押し付けられている。
【００１２】
操舵補助力発生用油圧アクチュエータとして油圧シリンダ２０が設けられている。この油圧シリンダ２０は、ラックハウジング９により構成されるシリンダチューブと、ラック１６に一体化されるピストン２１を備えている。そのピストン２１により仕切られる油室２２、２３に操舵抵抗に応じて圧油を供給するため、ロータリー式油圧制御弁３０が設けられている。
【００１３】
その制御弁３０は、バルブハウジング７に相対回転可能に挿入される筒状の第１バルブ部材３１と、この第１バルブ部材３１に同軸中心に相対回転可能に挿入される第２バルブ部材３２とを備える。その第１バルブ部材３１は出力軸３にピン２９により同行回転するよう連結されている。その第２バルブ部材３２は入力軸２と一体的に成形され、入力軸２の外周部により第２バルブ部材３２が構成され、第２バルブ部材３２は入力軸２と同行回転する。よって、第１バルブ部材３１と第２バルブ部材３２は、操舵抵抗に応じ前記トーションバー６がねじれることで同軸中心に相対回転する。
【００１４】
そのバルブハウジング７に、ポンプ７０に接続される入口ポート３４と、前記油圧シリンダ２０の一方の油室２２に接続される第１ポート３７と、他方の油室２３に接続される第２ポート３８と、直接にタンク７１に接続される第１出口ポート３６と、後述の可変絞り弁６０を介しタンク７１に接続される第２出口ポート６１とが設けられている。各ポート３４、３６、３７、３８、６１は、その第１バルブ部材３１と第２バルブ部材３２との内外周間の流路を介し互いに接続されている。
【００１５】
すなわち、図３、図４に示すように、第１バルブ部材３１の内周に溝５０ａ、５０ｂ、５０ｃが、周方向に間隔をおいた１２箇所に形成されている。また、第２バルブ部材３２の外周に溝５１ａ、５１ｂ、５１ｃが、周方向に間隔をおいた１２箇所に形成されている。図４は実線により第２バルブ部材３２の展開図を示し、鎖線により第１バルブ部材３１に形成された溝５０ａ、５０ｂ、５０ｃを示す。第１バルブ部材３１に形成された溝５０ａ、５０ｂ、５０ｃの間に第２バルブ部材３２に形成された溝５１ａ、５１ｂ、５１ｃが位置する。
【００１６】
その第１バルブ部材３１の溝は、３つの右操舵用溝５０ａと、３つの左操舵用溝５０ｂと、６つの連絡用溝５０ｃとを構成する。その右操舵用溝５０ａは、第１バルブ部材３１に形成された流路５３と第１ポート３７とを介し油圧シリンダ２０の右操舵補助力発生用油室２２に接続され、互いに周方向に１２０°離れて配置される。その左操舵用溝５０ｂは、第１バルブ部材３１に形成された流路５４と第２ポート３８とを介し油圧シリンダ２０の左操舵補助力発生用油室２３に接続され、互いに周方向に１２０°離れて配置される。
【００１７】
その第２バルブ部材３２の溝は、６つの圧油供給用溝５１ａと、３つの第１圧油排出用溝５１ｂと、３つの第２圧油排出用溝５１ｃとを構成する。その圧油供給用溝５１ａは、第１バルブ部材３１に形成された圧油供給路５５と入口ポート３４とを介しポンプ７０に接続され、互いに周方向に６０°離れて配置される。その第１圧油排出用溝５１ｂは、入力軸２に形成された流路５２ａから入力軸２とトーションバー６との間を通り、入力軸２に形成された流路５２ｂ（図１参照）と第１出口ポート３６とを介しタンク７１に接続され、互いに周方向に１２０°離れて配置される。その第２圧油排出用溝５１ｃは、第１バルブ部材３１に形成された流路５９と第２出口ポート６１とを介し可変絞り弁６０に接続され、互いに周方向に１２０°離れて配置される。
【００１８】
各第１圧油排出用溝５１ｂは右操舵用溝５０ａと左操舵用溝５０ｂの間に配置され、各第２圧油排出用溝５１ｃは連絡用溝５０ｃの間に配置され、右操舵用溝５０ａと連絡用溝５０ｃとの間および左操舵用溝５０ｂと連絡用溝５０ｃとの間に圧油供給用溝５１ａは配置される。
【００１９】
その第１バルブ部材３１に形成された溝５０ａ、５０ｂ、５０ｃの軸方向に沿う縁と第２バルブ部材３２に形成された溝５１ａ、５１ｂ、５１ｃの軸方向に沿う縁との間が絞り部Ａ、Ａ′、Ｂ、Ｂ′、Ｃ、Ｃ′、Ｄ、Ｄ′を構成する。これにより、各絞り部Ａ、Ａ′、Ｂ、Ｂ′、Ｃ、Ｃ′、Ｄ、Ｄ′はポンプ７０とタンク７１と油圧シリンダ２０とを接続する油路２７に配置されている。
【００２０】
図５に示すように、その第２バルブ部材３２に形成された溝５１ａ、５１ｂ、５１ｃの軸方向に沿う縁は面取り部とされている。各面取り部の周方向幅は、各絞り部Ａ、Ａ′、Ｂ、Ｂ′、Ｃ、Ｃ′、Ｄ、Ｄ′を全閉するのに要する両バルブ部材の相対回転角度である閉鎖角度に応じて定められている。すなわち、その圧油供給用溝５１ａと連絡用溝５０ｃとの間の絞り部Ａ′、Ｃ′における圧油供給用溝５１ａの軸方向に沿う縁（図３において□で囲む）の面取り部の周方向幅をＷ、連絡用溝５０ｃと第２圧油排出用溝５１ｃとの間の絞り部Ｂ′、Ｄ′における第２圧油排出用溝５１ｃの軸方向に沿う縁（図３において△で囲む）の面取り部の周方向幅をＷ′、その他の絞り部Ａ、Ｂ、Ｃ、Ｄにおける第２バルブ部材３２に形成された溝の軸方向に沿う縁（図３において○で囲む）の面取り部の周方向幅をＷ″として、図４、図５に示すように、Ｗ＞Ｗ′＞Ｗ″とされている。操舵抵抗のない状態（図４、図５の状態）から各絞り部Ａ、Ａ′、Ｂ、Ｂ′、Ｃ、Ｃ′、Ｄ、Ｄ′を全閉するのに要する両バルブ部材３１、３２の相対回転角度（すなわち閉鎖角度）を互いに比較すると、絞り部Ａ′、Ｃ′の閉鎖角度θｒは絞り部Ｂ′、Ｄ′の閉鎖角度θｓよりも大きく、両閉鎖角度θｒ、θｓは、他の各絞り部Ａ、Ｂ、Ｃ、Ｄの閉鎖角度θｔよりも大きい。これにより、第１バルブ部材３１と第２バルブ部材３２との間の各絞り部は、複数の絞り部Ａ、Ｂ、Ｃ、Ｄからなる第１の組と、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄよりも閉鎖角度の大きな複数の絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′からなる第２の組とに組分けされる。また、第２の組に属する絞り部は、絞り部Ｂ′、Ｄ′と、この絞り部Ｂ′、Ｄ′よりも閉鎖角度の大きな絞り部Ａ′、Ｃ′の２種類とされる。
【００２１】
その入力軸２と出力軸３は、路面から操舵用車輪を介し伝達される操舵抵抗によるトーションバー６のねじれによって相対回転する。その相対回転により第１バルブ部材３１と第２バルブ部材３２とが相対回転することで、各絞り部Ａ、Ｂ、Ｃ、Ｄ、Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積すなわち開度が変化し、油圧シリンダ２０が操舵抵抗に応じた操舵補助力を発生する。
【００２２】
すなわち、図４は操舵が行なわれていない状態を示し、両バルブ部材３１、３２の間の絞り部Ａ、Ｂ、Ｃ、Ｄ、Ａ′、Ｂ′、Ｃ′、Ｄ′は全て開かれ、入口ポート３４と各出口ポート３６、６１とは弁間流路２７を介し連通し、ポンプ７０から制御バルブ３０に流入する油はタンク７１に還流し、操舵補助力は発生しない。
【００２３】
この状態から右方へ操舵することによって生じる操舵抵抗により両バルブ部材３１、３２が相対回転すると、図３に示すように、絞り部Ａ、Ａ′の開度が大きくなり、絞り部Ｂ、Ｂ′の開度が小さくなり、絞り部Ｃ、Ｃ′の開度が小さくなり、絞り部Ｄ、Ｄ′の開度が大きくなる。これにより、図中矢印で示す圧油の流れにより油圧シリンダ２０の右操舵補助力発生用油室２２に操舵抵抗に応じた圧力の圧油が供給され、また、左操舵補助力発生用油室２３からタンク７１に油が還流し、車両の右方への操向補助力が油圧シリンダ２０からラック１６に作用する。
【００２４】
左方へ操舵すると第１バルブ部材３１と第２バルブ部材３２とが右方に操舵した場合と逆方向に相対回転し、絞り部Ａ、Ａ′の開度が小さくなり、絞り部Ｂ、Ｂ′の開度が大きくなり、絞り部Ｃ、Ｃ′の開度が大きくなり、絞り部Ｄ、Ｄ′の開度が小さくなるので、車両の左方への操舵補助力が油圧シリンダ２０からラック１６に作用する。
【００２５】
図１、図６に示すように、その第２出口ポート６１に連通する可変絞り弁６０は、バルブハウジング７に接続される第２バルブハウジング７′と、この第２バルブハウジング７′に形成された挿入孔６６に軸方向（図１、図６において上下方向）に移動可能に挿入されたスプール６２と、そのスプール６２にねじ合わされるネジ部材６４とを有する。その挿入孔６６の一端はプラグ６８により閉鎖され、他端はカバー９４′により閉鎖されている。そのスプール６２とプラグ６８との間に圧縮コイルバネ９０が配置されている。そのネジ部材６４にステッピングモータ８０が接続され、そのステッピングモータ８０にコントローラ（図示省略）が接続される。そのコントローラは車速センサ（図示省略）に接続され、そのステッピングモータ８０を車速に応じ制御する。すなわち、高速になるとネジ部材６４は一方向に回転してスプール６２は図中上方に変位し、低速になるとネジ部材６４は他方向に回転してスプール６２は図中下方に変位する。
【００２６】
そのスプール６２の外周に周溝６２ａが形成され、その挿入孔６６の内周に周溝６６ａが形成され、両周溝６２ａ、６６ａの間が可変絞り部６７とされている。その可変絞り部６７の開度は、高速になってスプール６２が図中上方に変位すると大きくなり、低速になってスプール６２が下方に変位すると小さくなる。
【００２７】
その挿入孔６６の内周の周溝６６ａと第２出口ポート６１とを連通する連絡流路５８が、スプール６２の径方向外方において第２バルブハウジング７′に形成されている。そのスプール６２の外周の周溝６２ａとスプール６２の通孔６２ｄとを連通する径方向孔６２ｃがスプール６２に形成されている。そのスプール６２の通孔６２ｄは、その挿入孔６６におけるスプール６２の下方空間に連絡する。そのスプール６２の下方空間と第１出口ポート３６とを連通する連絡流路７６が、スプール６２の径方向外方においてバルブハウジング７と第２バルブハウジング７′とに亘り形成されている。
【００２８】
これにより、ポンプ７０から供給される圧油は、前記弁間流路２７および第２出口ポート６１から連絡流路５８に導かれ、この連絡流路５８から可変絞り部６７に至り、この可変絞り部６７から連絡流路７６、第１出口ポート３６を介しタンク７１に至る。なお、スプール６２には通孔６２ｄと平行にドレン流路６２ｈが形成され、スプール６２の上方空間と下方空間とを接続する。
【００２９】
その可変絞り部６７の開度に対応する流路面積の最大値は、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の開度に対応する流路面積の最大値（両バルブ部材３１、３２の相対回転角が大きくなる程に流路面積が小さくなる特性における最大値である。すなわち、右操舵時は絞り部Ｂ′、Ｃ′の合計流路面積の最大値をいい、左操舵時は絞り部Ａ′、Ｄ′の合計流路面積の最大値をいう。以下「流路面積の最大値」という場合は同旨）以上、若しくは絞り機能を奏さなくなるまで大きくされている。その可変絞り部６７の流路面積の最小値は、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積の最小値（両バルブ部材３１、３２の相対回転角が大きくなる程に流路面積が小さくなる特性における最小値である。すなわち、右操舵時は絞り部Ｂ′、Ｃ′の合計流路面積の最小値をいい、左操舵時は絞り部Ａ′、Ｄ′の合計流路面積の最小値をいい、全閉状態を含む。以下「流路面積の最小値」という場合は同旨）以下とされる。
【００３０】
これにより、図２に示す油圧回路が構成され、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′とタンク７１との間の油路に、車速に応じて開度が変化する可変絞り部６７が可変絞り弁６０により設けられる。
【００３１】
図７において、実線Ｘは両バルブ部材３１、３２の相対回転角に対する第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄの開度に対応する流路面積の変化特性（その相対回転角が大きくなる程に流路面積が小さくなる特性である。この場合、右操舵時は絞り部Ｂ、Ｃの合計流路面積の変化特性をいい、左操舵時は絞り部Ａ、Ｄの合計流路面積の変化特性をいう。以下「流路面積の変化特性」という場合は同旨）を示す。１点鎖線Ｕは、その相対回転角に対する第２の組に属する圧油供給用溝５１ａと連絡用溝５０ｃとの間の絞り部Ａ′、Ｃ′の流路面積の変化特性を示す。１点鎖線Ｖは、その相対回転角に対する第２の組に属する連絡用溝５０ｃと第２圧油排出用溝５１ｃとの間の絞り部Ｂ′、Ｄ′の流路面積の変化特性を示す。実線Ｙは、その絞り部Ａ′、Ｃ′の流路面積の変化特性と絞り部Ｂ′、Ｄ′の流路面積の変化特性を合成した特性を示す。破線Ｒは可変絞り部６７の中速走行時における流路面積を示す。
【００３２】
低速走行時においては、スプール６２は図１、図６において下方に変位し、このスプール６２の変位により可変絞り部６７は全閉状態になる。よって、油圧シリンダ２０に作用する油圧は、第１の組の絞り部Ａ、Ｂ、Ｃ、Ｄの流路面積の変化特性線Ｘに応じ制御される。この場合、図８において実線αで示すように、操舵抵抗に対応する操舵トルクが小さく、両バルブ部材３１、３２の相対回転角が小さくても、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄの開度は小さいので、操舵トルクの変化に対して油圧変化が少ない領域を小さくし、操舵の高応答性を満足させて旋回性能を向上できる。
【００３３】
高速走行時においては、スプール６２は図１、図６において上方に変位し、このスプール６２の変位によって可変絞り部６７の流路面積は、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積の最大値以上になる。よって、油圧シリンダ２０に作用する油圧は、第２の組の絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積の変化特性線Ｙ及び第１の組の絞り部Ａ、Ｂ、Ｃ、Ｄの流路面積の変化特性線Ｘの合成特性に応じ制御される。この場合、図８において実線βで示すように、操舵トルクが大きく、両バルブ部材３１、３２の相対回転角が大きくても、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の開度は大きいので、操舵トルクの変化に対して油圧変化が少ない領域を大きくし、高速走行時における操舵の安定性を満足させることができる。
【００３４】
中速走行時においては、スプール６２の変位により可変絞り部６７の流路面積は、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積の最小値よりも大きく最大値よりも小さくなる。これにより、図７に示すように、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄの流路面積が最小値（本実施形態では全閉状態）になるまでの間（図７において両バルブ部材の相対回転角がθａになるまでの間）は、その第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄの流路面積の変化特性線Ｘに可変絞り部６７の流路面積の特性線Ｒを合成した特性に応じて、油圧シリンダ２０に作用する油圧が制御される。第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄが全閉状態になった時点から、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積が可変絞り部６７の流路面積よりも小さくなるまでの間（図７において両バルブ部材の相対回転角がθａとθｂとの間）では、可変絞り部６７の流路面積により定まる一定値になり、油圧シリンダ２０に作用する油圧は操舵抵抗に応じて制御できない。しかる後に、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積が可変絞り部６７の流路面積よりも小さくなると、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積の変化特性線Ｙに応じた操舵補助力が付与される。この場合、図８において実線γで示すように、操舵トルクの変化に対する油圧変化は、低速走行時と高速走行時の中間の特性を示す。
【００３５】
その第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄが全閉状態になった後に、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積が可変絞り部６７の流路面積よりも小さくなるまでの間（θａ〜θｂの間）は、その第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′が全閉状態になる点と、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄが全閉状態になる点との差（θｃ−θａ）を小さくすることなく、小さくされている。すなわち、絞り部Ｂ′、Ｄ′が絞り部Ａ′、Ｃ′と同様に図中１点鎖線Ｕで示す相対回転角に対する流路面積変化特性を有すると仮定すると、相対回転角に対する第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積の変化特性は、図７において２点鎖線Ｍで示すものになる。そうすると、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積が可変絞り部６７の流路面積よりも小さくなるまでの間（両バルブ部材の相対回転角がθａとθｄとの間）は大きくなるので、操舵補助力を操舵抵抗に応じ制御できない領域が大きくなる。これに対し、上記実施形態では、絞り部Ｂ′、Ｄ′の閉鎖角度θｓは絞り部Ａ′、Ｃ′の閉鎖角度θｒよりも小さいので、中速走行時において操舵補助力を操舵抵抗に応じ制御できない領域を小さくできる。しかも、絞り部Ｂ′、Ｄ′が全閉状態になる点（図７において両バルブ部材の相対回転角がθｅの点）では、絞り部Ａ′、Ｃ′は未だ閉じていないので、操舵補助力を操舵抵抗に応じ制御できる領域は小さくなることはない。
【００３６】
図９は、その第２バルブ部材３２の軸方向視の要部を示し、その第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄそれぞれにおける上記面取り部は、周方向に並列する複数（本実施形態では２つ）の平坦面Ｓ１、Ｓ２から構成されている。その第２バルブ部材３２の外接円の径方向に対して、前記各平坦面Ｓ１、Ｓ２に直角な方向の傾き角Φ１、Φ２は互いに相異するものとされている。その第２バルブ部材３２の溝５１ａ、５１ｂに近い方の平坦面Ｓ１に直角な方向の上記傾き角Φ１は、その溝５１ａ、５１ｂから遠い方の平坦面Ｓ２に直角な方向の上記傾き角Φ２よりも大きくされている。なお、図９において破線は面取り部の形成前の状態を示す。
【００３７】
上記構成によれば、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄそれぞれにおける上記面取り部を構成する２つの平坦面Ｓ１、Ｓ２の中で、第２バルブ部材３２の溝５１ａ、５１ｂから遠い方の平坦面Ｓ２に直角な方向の上記傾き角Φ２を小さくすることで、その第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄにおける絞り機能を確保しつつ面取り部の周方向幅Ｗ″を大きくできる。これにより、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄが全閉状態になった後に、第２の組に属する絞り部Ａ′、Ｂ′、Ｃ′、Ｄ′の流路面積が可変絞り部６７の流路面積よりも小さくなるまでの間（図７においてθａ〜θｂの間）を小さくできる。よって、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄの閉鎖タイミングを遅らせ、中速走行時において操舵補助力を操舵抵抗に応じ制御できない領域を小さくし、操舵トルク変化に対する操舵補助力発生用の油圧の追従性を向上できる。
【００３８】
しかも、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄそれぞれにおける上記面取り部を構成する２つの平坦面Ｓ１、Ｓ２の中で、第２バルブ部材３２の溝５１ａ、５１ｂに近い方の平坦面Ｓ１に直角な方向の上記傾き角Φ１を大きくすることで、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄにおいて閉鎖開始初期の流路面積が急激に小さくなるのを防止できる。すなわち、図９において２点鎖線で示すように、仮に第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄを第２バルブ部材３２の溝５１ａ、５１ｂから遠い方の平坦面Ｓ２のみで構成した場合、その絞り部Ａ、Ｂ、Ｃ、Ｄの閉鎖開始初期の平坦面Ｓ２と第１バルブ部材３１の溝５０ａ、５０ｂの縁との距離Ｌ１が小さくなる。そうすると、低中速走行時の操舵開始初期に操舵補助力発生用油圧が急上昇し、図８において破線α′、γ′で示すように、操舵トルクに対する操舵補助力発生用油圧の変化割合が急変するために操舵フィーリングが低下する。また、中速走行時には破線γ′で示すように、操舵補助力発生用油圧の実用域（図９において油圧Ｐａ以下）で、操舵トルクが小さくても操舵補助力発生用油圧が大きくなることで走行安定性が低下する。さらに、図８において破線β′で示すように、高速走行時に操舵補助力発生用油圧を十分に大きくするのに必要な操舵トルクが過大になり、必要な操舵補助を行うことができない。これに対して、上記構成によれば、第１の組に属する絞り部Ａ、Ｂ、Ｃ、Ｄの閉鎖開始初期の平坦面Ｓ１と第１バルブ部材３１の溝５０ａ、５０ｂ、５０ｃの縁との距離を図９においてＬ２だけ大きくできる。よって、低中速走行時の操舵開始初期に操舵補助力発生用油圧が急上昇することはなく、操舵トルクに対して略一定割合で操舵補助力発生用油圧を増加させて操舵フィーリングを向上できる。また、中速走行時において、操舵補助力発生用油圧が過大になるのを防止して走行安定性を向上できる。また、高速走行時に操舵補助力発生用油圧を十分に大きくするのに必要な操舵トルクが過大になることはなく、必要な操舵補助を行うことができる。
【００３９】
図１０は上記ステッピングモータ８０の第１変形例を示す。上記実施形態ではモータ８０の回転を減速することなくネジ部材６４に伝達しているが、この第１変形例ではトラクションドライブ型減速機構１０１により減速している。すなわち、そのモータ８０の出力シャフト８０ａはネジ部材６４にブッシュ１０９を介して相対回転可能に挿入される。その減速機構１０１は、そのネジ部材６４に一体化される支持シャフト１０２と、この支持シャフト１０２に回転可能に嵌め合わされる遊星ローラ１０３と、第２バルブハウジング７′に一体化されるリング１０４とを有する。その遊星ローラ１０３が出力シャフト８０ａの外周とリング１０４の内周とに接触する。これにより、その出力シャフト８０ａの回転が摩擦により遊星ローラ１０３に伝達され、その遊星ローラ１０３がリング１０４との間の摩擦により出力シャフト８０ａの周りを周回することでネジ部材６４が回転駆動される。これにより、モータ８０を大型化することなく、出力シャフト８０ａとネジ部材６４とを同軸心に配置したままで、平歯車を用いた減速機構に比べてコンパクトな構成で、ギヤを用いる場合のようなバックラッシによる回転伝達ロスなしに、モータ８０の出力トルクを増加したのと同様の効果を得られ、高低温時における作動を確実に行える。他は上記実施形態と同様である。
【００４０】
図１１は上記ステッピングモータ８０の第２変形例を示す。上記実施形態ではモータ８０の回転を減速することなくネジ部材６４に伝達しているが、この第２変形例では遊星歯車機構２０１により減速している。すなわち、そのモータ８０の出力シャフト８０ａはネジ部材６４にブッシュ２０９を介して相対回転可能に挿入される。その遊星歯車機構２０１は、そのネジ部材６４に一体化される支持シャフト２０２と、この支持シャフト２０２に回転可能に嵌め合わされる遊星ギア２０３と、第２バルブハウジング７′に一体化されるリングギア２０４と、その出力シャフト８０ａに一体化されるサンギア２０５とを有する。その遊星ギア２０３がリングギア２０４とサンギア２０５とに噛み合う。これにより、その出力シャフト８０ａの回転がサンギア２０５から遊星ギア２０３に伝達され、その遊星ギア２０３がリングギア２０４との噛み合いにより出力シャフト８０ａの周りを周回することでネジ部材６４が回転駆動される。これにより、モータ８０を大型化することなく、出力シャフト８０ａとネジ部材６４とを同軸心に配置したままで、平歯車を用いた減速機構に比べてコンパクトな構成で、モータ８０の出力トルクを増加したのと同様の効果を得られ、高低温時における作動を確実に行える。他は上記実施形態と同様である。
【００４１】
本発明は上記実施形態に限定されない。例えば、上記実施形態では第１の組に属する絞り部における面取り部は、周方向に並列する２つの平坦面から構成したが、３つ以上の平坦面から構成してもよい。また、上記実施形態では第２の組に属する絞り部の閉鎖角度は互いに異なるものとしたが、同一であってもよい。また、各バルブ部材における溝の数は限定されない。
【００４２】
【発明の効果】
本発明によれば、操舵フィーリング、走行安定性を向上し、高速走行時でも必要な操舵補助を行うことができる油圧パワーステアリング装置を提供できる。
【図面の簡単な説明】
【図１】本発明の実施形態の油圧パワーステアリング装置の縦断面図
【図２】本発明の実施形態の油圧パワーステアリング装置の油圧回路を示す図
【図３】本発明の実施形態の油圧パワーステアリング装置における制御弁の横断面構造の説明図
【図４】本発明の実施形態の油圧パワーステアリング装置の制御弁の展開図
【図５】本発明の実施形態の油圧パワーステアリング装置の制御弁の部分拡大図
【図６】本発明の実施形態の油圧パワーステアリング装置の可変絞り弁の縦断面図
【図７】本発明の実施形態の油圧パワーステアリング装置における制御弁の絞り部の開度とバルブ部材の相対回転角との関係を示す図
【図８】油圧パワーステアリング装置における操舵トルクと油圧との関係を示す図
【図９】本発明の実施形態の油圧パワーステアリング装置の制御弁の要部の拡大図
【図１０】本発明の第１変形例のモータの部分断面図
【図１１】本発明の第２変形例のモータの部分断面図
【符号の説明】
２０油圧シリンダ
３０制御弁
３１第１バルブ部材
３２第２バルブ部材
５０ａ、５０ｂ、５０ｃ、５１ａ、５１ｂ、５１ｃ溝
６７可変絞り部
７０ポンプ
７１タンク
Ａ、Ｂ、Ｃ、Ｄ第１の組に属する絞り部
Ａ′、Ｂ′、Ｃ′、Ｄ′ 第２の組に属する絞り部
Ｓ１、Ｓ２平坦面[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a rack and pinion hydraulic power steering device that controls a hydraulic pressure acting on a hydraulic actuator for generating a steering assist force by a rotary type control valve.
[0002]
[Prior art]
In the hydraulic power steering apparatus, the steering assist force is generated by controlling the hydraulic pressure acting on the hydraulic actuator with a control valve. As the control valve, a rotary type valve having a cylindrical first valve member and a second valve member inserted into the first valve member so as to be relatively rotatable in accordance with a steering resistance is used.
[0003]
In the rotary type control valve, a plurality of axial grooves are formed on the inner periphery of the first valve member at intervals in the circumferential direction, and the outer periphery of the second valve member is along the plurality of axial directions. Grooves are formed at intervals in the circumferential direction. A plurality of throttles whose opening varies between the edge along the axial direction of the groove of the first valve member and the edge along the axial direction of the groove of the second valve member according to the relative rotation angle of both valve members. It is considered to be a part. By connecting the actuator to the pump and the tank via the control valve, it is possible to apply a steering assist force according to the change in the opening of each throttle.
[0004]
In addition, in order to improve the stability of the vehicle during high speed travel and the turning performance during low speed travel, the steering assist force is reduced during high speed travel and the steering assist force is increased during low speed travel. For this reason, a configuration is adopted in which the steering assist force is changed according to the driving conditions. That is, each throttle part is divided into a first group and a second group, the closing angle of the throttle part belonging to the second group is made larger than the closing angle of the throttle part belonging to the first group, A variable throttle portion whose opening degree changes according to operating conditions is provided in the oil passage between the throttle portion and the tank belonging to the second set. The opening of the variable throttle is increased during high-speed travel and decreased during low-speed travel. As a result, the hydraulic pressure that generates the steering assist force changes according to the opening degree of the throttle portions of both the first group and the second group during high-speed traveling, and the second pressure decreases as the speed decreases. The influence of the opening of the throttle part of the set is reduced. As a result, the region where the hydraulic pressure change is small with respect to the steering torque change corresponding to the steering resistance is large during high-speed traveling, and the region where the hydraulic pressure change is small with respect to the steering torque change is small during low-speed traveling.
[0005]
In order to set the closing angle of each throttle part, the edge along the axial direction of each groove in the second valve member is a chamfered part, and the circumferential width of this chamfered part is used to fully close each throttle part. It is determined according to the closing angle that is the relative rotation angle of both valve members.
[0006]
[Problems to be solved by the invention]
In the medium speed traveling state in which the variable throttle unit is opened, the aperture of the throttle unit belonging to the second group is smaller than the aperture of the variable throttle unit after the throttle unit belonging to the first group is closed. In the meantime, since the opening of the variable throttle is constant, the hydraulic pressure acting on the hydraulic actuator cannot be controlled according to the steering resistance. That is, during that time, the steering assist force generation hydraulic pressure does not follow the steering torque change, and the steering feeling is lowered.
[0007]
Therefore, by closing the closing timing of the throttle part belonging to the first group as much as possible, the opening degree of the throttle part belonging to the second group is opened after the throttle part belonging to the first group is closed. It is conceivable to reduce the time until it becomes smaller than the degree.
[0008]
However, each of the conventional chamfered portions is configured by a single flat surface. Therefore, in order to delay the closing timing of the throttle parts belonging to the first group, it is necessary to increase the circumferential width of the chamfered part of the throttle part belonging to the first group. In this case, in order to secure a stop function with respect to the radial direction of the circumscribed circle of the second valves members, reduce the perpendicular direction of the tilt angle to the flat surfaces constituting the chamfer of the diaphragm portion of the first set There is a need to. As a result, the flow passage area at the beginning of closing of the throttle portion belonging to the first group is rapidly reduced. For this reason, the steering assist force generating hydraulic pressure suddenly rises at the beginning of steering during low and medium speed traveling, and the change rate of the steering assist force generating hydraulic pressure with respect to the steering torque changes suddenly, so that the steering feeling decreases. Further, even when the steering torque is small during medium speed traveling, the traveling assisting force generating hydraulic pressure increases, so that traveling stability decreases. Furthermore, there is a problem that the steering torque necessary for sufficiently increasing the hydraulic pressure for generating the steering assist force during high speed traveling becomes excessive, and the necessary steering assist cannot be performed.
[0009]
An object of the present invention is to provide a hydraulic power steering device that can solve the above-described problems.
[0010]
[Means for Solving the Problems]
The present invention includes a steering assist force generating hydraulic actuator and a hydraulic control valve acting on the actuator. The control valve has a cylindrical first valve member and the first valve member in accordance with the steering resistance. A second valve member inserted so as to be relatively rotatable, and a plurality of axially extending grooves are formed on the inner periphery of the first valve member at intervals in the circumferential direction. A plurality of grooves along the axial direction are formed on the outer periphery of the first valve member at intervals in the circumferential direction, an edge along the axial direction of the groove of the first valve member, and an edge along the axial direction of the groove of the second valve member; A plurality of throttle parts whose opening degree changes in accordance with the relative rotation angle of both valve members are provided between the two valve members, and through the control valves so that a steering assist force can be applied according to the opening degree change of each throttle part. The actuator is connected to a pump and a tank, The edge along the axial direction of each groove in the two-valve member is a chamfered portion, and the circumferential width of this chamfered portion corresponds to the closing angle that is the relative rotation angle of both valve members required to fully close each throttle portion. Each throttle part is divided into a first group and a second group, and the closing angle of the throttle part belonging to the second group is made larger than the closing angle of the throttle part belonging to the first group. The present invention is applied to a hydraulic power steering apparatus in which an oil passage between a throttle portion and a tank belonging to the second group is provided with a variable throttle portion whose opening degree changes according to operating conditions.
In the hydraulic power steering apparatus of the present invention, the chamfered portion of the throttle section belonging to the first set includes a plurality of flat surfaces parallel to the circumferential direction, the radial direction of the circumscribed circle of the second valves members On the other hand, the inclination angles in the direction perpendicular to the flat surfaces are different from each other, and the direction perpendicular to the flat surface closer to the groove of the second valve member is the flat surface far from the groove. The tilt angle is larger than the direction perpendicular to the angle.
According to the configuration of the present invention, the circumscribed circle in a direction perpendicular to the flat surface far from the groove of the second valve member among the plurality of flat surfaces constituting the chamfered portion of the throttle portion belonging to the first group. By reducing the inclination angle with respect to the radial direction, the circumferential width of the chamfered portion can be increased while ensuring the diaphragm function in the diaphragm portion belonging to the first group. Thereby, after the flow passage area of the throttle part belonging to the first group is minimized, the time until the opening degree of the throttle part belonging to the second group becomes smaller than the opening degree of the variable throttle part can be reduced. . Therefore, the closing timing of the throttle parts belonging to the first group is delayed, the region in which the steering assist force cannot be controlled according to the steering resistance during medium speed traveling is reduced, and the followability of the hydraulic pressure for generating the steering assist force with respect to the steering torque change is reduced. Can be improved.
Moreover, an inclination angle with respect to the radial direction of the circumscribed circle in a direction perpendicular to the flat surface closer to the groove of the second valve member among the plurality of flat surfaces constituting the chamfered portion in the throttle portion belonging to the first group. By enlarging, it is possible to prevent the flow channel area at the beginning of closing from rapidly decreasing in the throttle portions belonging to the first group. As a result, the steering assist force generation hydraulic pressure does not rise rapidly at the beginning of steering at low and medium speeds, and the steering assist force generation hydraulic pressure is increased at a substantially constant rate with respect to the steering torque to improve the steering feeling. it can. In addition, during running at medium speed, it is possible to prevent the steering assist force generation hydraulic pressure from becoming excessive and improve running stability. Furthermore, the steering torque necessary to sufficiently increase the hydraulic pressure for generating the steering assist force during high-speed traveling does not become excessive, and the necessary steering assist can be performed.
[0011]
DETAILED DESCRIPTION OF THE INVENTION
A rack and pinion hydraulic power steering apparatus 1 according to an embodiment of the present invention shown in FIG. 1 includes an input shaft 2 connected to a vehicle handle (not shown) and an output connected to the input shaft 2 via a torsion bar 6. A shaft 3 is provided. The torsion bar 6 is connected to the input shaft 2 by a pin 4 and is connected to the output shaft 3 by a serration 5. The input shaft 2 is supported by the valve housing 7 via a bearing 8 and supported by the output shaft 3 via a bush 12. The output shaft 3 is supported by the rack housing 9 via bearings 10 and 11. A pinion 15 is formed on the output shaft 3, and a steering wheel (not shown) is connected to a rack 16 that meshes with the pinion 15. Thereby, the rotation of the input shaft 2 by the steering is transmitted to the pinion 15 through the torsion bar 6, and the rack 16 moves in the vehicle width direction by the rotation of the pinion 15, and the steering of the vehicle is performed by the movement of the rack 16. Made. Oil seals 42 and 43 are interposed between the input / output shafts 2 and 3 and the housing 7. A support yoke 40 that supports the rack 16 is pressed against the rack 16 by the elasticity of the spring 41.
[0012]
A hydraulic cylinder 20 is provided as a steering assist force generating hydraulic actuator. The hydraulic cylinder 20 includes a cylinder tube constituted by the rack housing 9 and a piston 21 integrated with the rack 16. In order to supply pressure oil to the oil chambers 22 and 23 partitioned by the piston 21 according to the steering resistance, a rotary hydraulic control valve 30 is provided.
[0013]
The control valve 30 includes a cylindrical first valve member 31 that is inserted into the valve housing 7 so as to be relatively rotatable, and a second valve member 32 that is inserted into the first valve member 31 so as to be relatively rotatable about a coaxial center. Is provided. The first valve member 31 is connected to the output shaft 3 by a pin 29 so as to rotate together. The second valve member 32 is formed integrally with the input shaft 2, and the second valve member 32 is configured by the outer peripheral portion of the input shaft 2, and the second valve member 32 rotates along with the input shaft 2. Therefore, the first valve member 31 and the second valve member 32 rotate relative to each other about the coaxial center by twisting the torsion bar 6 according to the steering resistance.
[0014]
The valve housing 7 has an inlet port 34 connected to the pump 70, a first port 37 connected to one oil chamber 22 of the hydraulic cylinder 20, and a second port 38 connected to the other oil chamber 23. And a first outlet port 36 directly connected to the tank 71 and a second outlet port 61 connected to the tank 71 via a variable throttle valve 60 described later. Each port 34, 36, 37, 38, 61 is connected to each other via a flow path between the inner and outer circumferences of the first valve member 31 and the second valve member 32.
[0015]
That is, as shown in FIGS. 3 and 4, grooves 50 a, 50 b, and 50 c are formed in the inner periphery of the first valve member 31 at twelve locations spaced in the circumferential direction. Further, grooves 51 a, 51 b, 51 c are formed on the outer periphery of the second valve member 32 at twelve locations spaced in the circumferential direction. FIG. 4 shows a developed view of the second valve member 32 by a solid line, and shows grooves 50a, 50b, 50c formed in the first valve member 31 by a chain line. The grooves 51a, 51b, 51c formed in the second valve member 32 are positioned between the grooves 50a, 50b, 50c formed in the first valve member 31.
[0016]
The grooves of the first valve member 31 constitute three right steering grooves 50a, three left steering grooves 50b, and six communication grooves 50c. The right steering groove 50a is connected to the right steering assist force generating oil chamber 22 of the hydraulic cylinder 20 via a flow path 53 formed in the first valve member 31 and the first port 37, and 120 in the circumferential direction. ° Located apart. The left steering groove 50b is connected to the left steering assist force generating oil chamber 23 of the hydraulic cylinder 20 via a flow path 54 formed in the first valve member 31 and the second port 38, and 120 in the circumferential direction. ° Located apart.
[0017]
The grooves of the second valve member 32 constitute six pressure oil supply grooves 51a, three first pressure oil discharge grooves 51b, and three second pressure oil discharge grooves 51c. The pressure oil supply groove 51a is connected to the pump 70 via the pressure oil supply path 55 formed in the first valve member 31 and the inlet port 34, and is arranged 60 ° apart from each other in the circumferential direction. The first pressure oil discharging groove 51b passes between the input shaft 2 and the torsion bar 6 from the flow path 52a formed in the input shaft 2, and the flow path 52b formed in the input shaft 2 (see FIG. 1). And the first outlet port 36 are connected to the tank 71 and are arranged 120 ° apart from each other in the circumferential direction. The second pressure oil discharging groove 51c is connected to the variable throttle valve 60 via a flow path 59 formed in the first valve member 31 and the second outlet port 61, and is disposed 120 ° apart in the circumferential direction. The
[0018]
Each of the first pressure oil discharge grooves 51b is disposed between the right steering groove 50a and the left steering groove 50b, and each of the second pressure oil discharge grooves 51c is disposed between the communication grooves 50c. The pressure oil supply groove 51a is disposed between the groove 50a and the communication groove 50c and between the left steering groove 50b and the communication groove 50c.
[0019]
A throttle portion is formed between the edges along the axial direction of the grooves 50a, 50b, and 50c formed in the first valve member 31 and the edges along the axial direction of the grooves 51a, 51b, and 51c formed in the second valve member 32. A, A ′, B, B ′, C, C ′, D, D ′ are formed. Accordingly, the throttle portions A, A ′, B, B ′, C, C ′, D, and D ′ are arranged in the oil passage 27 that connects the pump 70, the tank 71, and the hydraulic cylinder 20.
[0020]
As shown in FIG. 5, the edges along the axial direction of the grooves 51a, 51b, 51c formed in the second valve member 32 are chamfered portions. The circumferential width of each chamfered portion is a closing angle that is a relative rotation angle of both valve members required to fully close each throttle portion A, A ′, B, B ′, C, C ′, D, D ′. It is determined accordingly. That is, the chamfered portion of the chamfered portion (enclosed by □ in FIG. 3) along the axial direction of the pressure oil supply groove 51a in the throttle portions A ′ and C ′ between the pressure oil supply groove 51a and the communication groove 50c. The circumferential width is W, and the edge along the axial direction of the second pressure oil discharge groove 51c at the narrowed portions B ′ and D ′ between the communication groove 50c and the second pressure oil discharge groove 51c (Δ in FIG. 3) The circumferential width of the chamfered portion is surrounded by W ′, and the edge along the axial direction of the groove formed in the second valve member 32 in the other throttle portions A, B, C, D (circled in FIG. 3) As shown in FIG. 4 and FIG. 5, W> W ′> W ″, where W ″ is the circumferential width of the chamfered portion. Both valve members 31, 32 required to fully close the throttle portions A, A ', B, B', C, C ', D, D' from the state without steering resistance (the state shown in FIGS. 4 and 5). Are compared with each other, the closing angle θr of the throttle parts A ′ and C ′ is larger than the closing angle θs of the throttle parts B ′ and D ′, and both the closing angles θr and θs are different from each other. Is larger than the closing angle θt of each of the aperture portions A, B, C, and D. Thereby, each throttle part between the 1st valve member 31 and the 2nd valve member 32 is the 1st group which consists of a plurality of throttle parts A, B, C, and D, and the throttle part which belongs to the 1st group. They are grouped into a second group consisting of a plurality of apertures A ′, B ′, C ′, D ′ having a larger closing angle than A, B, C, D. Further, there are two types of apertures belonging to the second group: apertures B ′ and D ′ and apertures A ′ and C ′ having a closing angle larger than that of the apertures B ′ and D ′.
[0021]
The input shaft 2 and the output shaft 3 rotate relative to each other by the torsion of the torsion bar 6 due to the steering resistance transmitted from the road surface via the steering wheel. By the relative rotation of the first valve member 31 and the second valve member 32 due to the relative rotation, the flow passage areas of the throttle portions A, B, C, D, A ′, B ′, C ′, D ′, that is, The opening degree changes, and the hydraulic cylinder 20 generates a steering assist force corresponding to the steering resistance.
[0022]
That is, FIG. 4 shows a state in which steering is not performed, and the throttle portions A, B, C, D, A ′, B ′, C ′, D ′ between the valve members 31, 32 are all opened, The inlet port 34 and each of the outlet ports 36 and 61 communicate with each other via the inter-valve flow path 27, and oil flowing into the control valve 30 from the pump 70 returns to the tank 71, and no steering assist force is generated.
[0023]
When the valve members 31 and 32 are rotated relative to each other by the steering resistance generated by steering to the right from this state, as shown in FIG. 3, the opening degree of the throttle portions A and A ′ increases, and the throttle portions B and B The opening of 'is reduced, the opening of the throttles C and C' is reduced, and the opening of the throttles D and D 'is increased. As a result, the pressure oil having a pressure corresponding to the steering resistance is supplied to the right steering assist force generating oil chamber 22 of the hydraulic cylinder 20 by the flow of the pressure oil indicated by the arrow in the drawing, and the left steering assist force generating oil chamber is provided. The oil flows back from 23 to the tank 71, and the steering assist force to the right of the vehicle acts on the rack 16 from the hydraulic cylinder 20.
[0024]
When steered to the left, the first valve member 31 and the second valve member 32 rotate relative to each other in the opposite direction to that when steered to the right, and the apertures of the throttle parts A and A ′ become smaller, so that the throttle parts B and B 'Is increased, the apertures of the throttles C and C' are increased, and the apertures of the throttles D and D 'are decreased, so that the steering assist force to the left of the vehicle is generated from the hydraulic cylinder 20 to the rack. 16 acts.
[0025]
As shown in FIGS. 1 and 6, the variable throttle valve 60 communicating with the second outlet port 61 is formed in the second valve housing 7 ′ connected to the valve housing 7 and the second valve housing 7 ′. The spool 62 is inserted into the insertion hole 66 so as to be movable in the axial direction (vertical direction in FIGS. 1 and 6), and the screw member 64 is screwed into the spool 62. One end of the insertion hole 66 is closed by a plug 68, and the other end is closed by a cover 94 '. A compression coil spring 90 is disposed between the spool 62 and the plug 68. A stepping motor 80 is connected to the screw member 64, and a controller (not shown) is connected to the stepping motor 80. The controller is connected to a vehicle speed sensor (not shown) and controls the stepping motor 80 according to the vehicle speed. That is, when the speed is high, the screw member 64 rotates in one direction and the spool 62 is displaced upward in the figure, and when the speed is low, the screw member 64 is rotated in the other direction and the spool 62 is displaced downward in the figure.
[0026]
A circumferential groove 62 a is formed on the outer periphery of the spool 62, a circumferential groove 66 a is formed on the inner periphery of the insertion hole 66, and a variable throttle portion 67 is formed between both the circumferential grooves 62 a and 66 a. The opening degree of the variable throttle 67 increases when the spool 62 is displaced upward in the drawing at a high speed, and decreases when the spool 62 is displaced downward at a low speed.
[0027]
A communication flow path 58 that communicates the inner circumferential groove 66 a of the insertion hole 66 with the second outlet port 61 is formed in the second valve housing 7 ′ on the radially outer side of the spool 62. A radial hole 62 c is formed in the spool 62 to communicate the circumferential groove 62 a on the outer periphery of the spool 62 and the through hole 62 d of the spool 62. The through hole 62 d of the spool 62 communicates with the space below the spool 62 in the insertion hole 66. A communication flow path 76 that communicates the space below the spool 62 and the first outlet port 36 is formed across the valve housing 7 and the second valve housing 7 ′ on the radially outer side of the spool 62.
[0028]
As a result, the pressure oil supplied from the pump 70 is guided from the inter-valve flow path 27 and the second outlet port 61 to the communication flow path 58, and reaches the variable throttle portion 67 from the communication flow path 58. From the part 67 to the tank 71 via the communication channel 76 and the first outlet port 36. The spool 62 is formed with a drain passage 62h parallel to the through hole 62d, and connects the upper space and the lower space of the spool 62.
[0029]
The maximum value of the channel area corresponding to the opening of the variable throttle 67 is the maximum of the channel area corresponding to the openings of the throttles A ′, B ′, C ′, D ′ belonging to the second group. (This is the maximum value in the characteristic that the flow path area decreases as the relative rotation angle of both valve members 31 and 32 increases. That is, the maximum value of the total flow path area of the throttle portions B 'and C' during right steering. During left steering, it refers to the maximum value of the total flow area of the throttles A 'and D' (hereinafter referred to as "maximum value of the flow area"), or increased until the throttle function is not achieved. ing. The minimum value of the flow passage area of the variable restricting portion 67 is the minimum value of the flow passage areas of the restricting portions A ′, B ′, C ′, D ′ belonging to the second group (relative rotation of both valve members 31, 32). The minimum value in the characteristic that the flow path area decreases as the angle increases, that is, the minimum value of the total flow path area of the throttle parts B ′ and C ′ during right steering, and the throttle part A during left steering. The minimum value of the total flow area of ′ and D ′, including the fully closed state (hereinafter referred to as “minimum value of the flow area” is the same).
[0030]
Accordingly, the hydraulic circuit shown in FIG. 2 is configured, and the opening degree is set in the oil passage between the throttle portions A ′, B ′, C ′, D ′ belonging to the second group and the tank 71 according to the vehicle speed. A variable throttle portion 67 that changes is provided by the variable throttle valve 60.
[0031]
In FIG. 7, a solid line X represents a change characteristic of the flow path area corresponding to the opening degree of the throttle portions A, B, C, and D belonging to the first group with respect to the relative rotation angle of both valve members 31 and 32 (its relative rotation angle). In this case, the change characteristic of the total flow area of the throttle parts B and C during the right steering, and the total flow of the throttle parts A and D during the left steering. This means the change characteristic of the road area (hereinafter referred to as “change characteristic of the flow path area”). An alternate long and short dash line U indicates a change characteristic of the flow path area of the throttle portions A ′ and C ′ between the pressure oil supply groove 51a and the communication groove 50c belonging to the second group with respect to the relative rotation angle. An alternate long and short dash line V indicates a change characteristic of the flow passage area of the throttle portions B ′ and D ′ between the communication groove 50c and the second pressure oil discharge groove 51c belonging to the second group with respect to the relative rotation angle. . A solid line Y indicates a characteristic obtained by synthesizing the change characteristics of the flow passage areas of the throttle portions A ′ and C ′ and the change characteristics of the flow passage areas of the throttle portions B ′ and D ′. A broken line R indicates the flow path area during the medium speed travel of the variable throttle portion 67.
[0032]
During low-speed travel, the spool 62 is displaced downward in FIGS. 1 and 6, and the variable restrictor 67 is fully closed by the displacement of the spool 62. Therefore, the hydraulic pressure acting on the hydraulic cylinder 20 is controlled according to the change characteristic line X of the flow path area of the first set of throttle portions A, B, C, and D. In this case, as shown by a solid line α in FIG. 8, even if the steering torque corresponding to the steering resistance is small and the relative rotation angles of both the valve members 31 and 32 are small, the throttle portions A, B, Since the opening degree of C and D is small, the region where the hydraulic pressure change is small with respect to the change of the steering torque can be reduced, and the turning performance can be improved by satisfying the high response of steering.
[0033]
During high-speed travel, the spool 62 is displaced upward in FIGS. 1 and 6, and the displacement of the spool 62 causes the flow area of the variable restrictor 67 to be the restrictors A ′, B ′, It becomes more than the maximum value of the channel area of C 'and D'. Therefore, the hydraulic pressure acting on the hydraulic cylinder 20 is the change characteristic line Y of the flow path area of the second set of throttle portions A ′, B ′, C ′, D ′ and the first set of throttle portions A, B, Control is performed according to the composite characteristic of the change characteristic line X of the flow path areas of C and D. In this case, as indicated by the solid line β in FIG. 8, even if the steering torque is large and the relative rotational angles of both valve members 31 and 32 are large, the throttle portions A ′, B ′, C ′, Since the opening degree of D ′ is large, the region where the change in hydraulic pressure is small with respect to the change in steering torque can be increased to satisfy the steering stability during high-speed running.
[0034]
During medium speed running, the flow path area of the variable throttle 67 is smaller than the minimum value of the flow areas of the throttles A ′, B ′, C ′, D ′ belonging to the second group due to the displacement of the spool 62. Greatly smaller than the maximum value. As a result, as shown in FIG. 7, the flow passage areas of the throttle portions A, B, C, and D belonging to the first group become the minimum value (in the present embodiment, the fully closed state) (in FIG. 7, Until the relative rotation angle of the two valve members reaches θa), the flow path of the variable throttle 67 is shown in the change characteristic line X of the flow areas of the throttles A, B, C, D belonging to the first group. The hydraulic pressure acting on the hydraulic cylinder 20 is controlled according to the characteristic obtained by combining the characteristic lines R of the area. The flow passage areas of the throttle parts A ′, B ′, C ′, D ′ belonging to the second group are variable from the time when the throttle parts A, B, C, D belonging to the first group are fully closed. Until it becomes smaller than the flow path area of the throttle part 67 (in FIG. 7, the relative rotation angles of both valve members are between θa and θb), it becomes a constant value determined by the flow path area of the variable throttle part 67, The hydraulic pressure acting on the hydraulic cylinder 20 cannot be controlled according to the steering resistance. Thereafter, when the flow passage areas of the throttle portions A ′, B ′, C ′, D ′ belonging to the second group become smaller than the flow passage area of the variable throttle portion 67, the throttle portions A ′ belonging to the second group. , B ′, C ′, and D ′ are applied with a steering assist force corresponding to the change characteristic line Y of the flow path area. In this case, as indicated by a solid line γ in FIG. 8, the change in hydraulic pressure with respect to the change in steering torque shows an intermediate characteristic between low speed running and high speed running.
[0035]
After the throttle portions A, B, C, and D belonging to the first group are fully closed, the flow passage areas of the throttle portions A ′, B ′, C ′, and D ′ belonging to the second group are variable. The throttle portions A ′, B ′, C ′, and D ′ belonging to the second set are fully closed until the flow passage area of the throttle portion 67 becomes smaller (between θa and θb). And the difference (θc−θa) from the point where the aperture portions A, B, C, and D belonging to the first group are in the fully closed state are made small. That is, if it is assumed that the throttle portions B ′ and D ′ have the flow path area change characteristic with respect to the relative rotation angle indicated by the one-dot chain line U in the drawing similarly to the throttle portions A ′ and C ′, the second relative to the relative rotation angle. The change characteristics of the flow path areas of the throttle portions A ′, B ′, C ′, and D ′ belonging to the set are shown by a two-dot chain line M in FIG. Then, until the flow passage area of the throttle portions A ′, B ′, C ′, D ′ belonging to the second group becomes smaller than the flow passage area of the variable throttle portion 67 (the relative rotation angle of both valve members is Since (between θa and θd) increases, the region in which the steering assist force cannot be controlled according to the steering resistance increases. On the other hand, in the above embodiment, the closing angle θs of the throttle parts B ′ and D ′ is smaller than the closing angle θr of the throttle parts A ′ and C ′, so that the steering assist force depends on the steering resistance during medium speed traveling. The area that cannot be controlled can be reduced. Moreover, at the point where the throttle parts B ′ and D ′ are fully closed (the relative rotation angle of both valve members in FIG. 7 is the point of θe), the throttle parts A ′ and C ′ are not yet closed. The region in which the force can be controlled according to the steering resistance is never reduced.
[0036]
FIG. 9 shows the main part of the second valve member 32 as viewed in the axial direction, and the chamfered parts in each of the throttle parts A, B, C, D belonging to the first group are plural in parallel in the circumferential direction ( In this embodiment, it is composed of two flat surfaces S1 and S2. The radial direction of the circumscribed circle of the second valves members 32, the tilt angle Φ1 of the direction perpendicular to the flat surfaces S1, S2, .phi.2 is intended to differences with each other. The inclination angle Φ1 in the direction perpendicular to the flat surface S1 closer to the grooves 51a and 51b of the second valve member 32 is the inclination angle Φ2 in the direction perpendicular to the flat surface S2 far from the grooves 51a and 51b. Has been bigger than. In FIG. 9, a broken line indicates a state before the chamfered portion is formed.
[0037]
According to the above-described configuration, the grooves 51a and 51b of the second valve member 32 among the two flat surfaces S1 and S2 constituting the chamfered portion in each of the throttle portions A, B, C, and D belonging to the first set. By reducing the inclination angle Φ2 in the direction perpendicular to the flat surface S2 far from the center, the circumferential direction of the chamfered portion is ensured while maintaining the aperture function in the aperture portions A, B, C, and D belonging to the first set. The width W ″ can be increased. As a result, after the throttle portions A, B, C, D belonging to the first group are fully closed, the throttle portions A ′, B ′, C ′ belonging to the second group are fully closed. , D ′ can be reduced until the flow area of the variable restrictor 67 becomes smaller than the flow area of the variable restrictor 67 (between θa and θb in FIG. 7), so that the restrictors A and B belonging to the first set. , Delay the closing timing of C and D, and adjust the steering assist force to the steering resistance when driving at medium speed. Uncontrolled area reduced, thereby improving the oil pressure of the follow-up of the steering assist force generated with respect to the steering torque change.
[0038]
Moreover, of the two flat surfaces S1 and S2 constituting the chamfered portion in each of the throttle portions A, B, C and D belonging to the first group, the one closer to the grooves 51a and 51b of the second valve member 32 by increasing the inclination angle Φ1 of the direction perpendicular to the flat surface S 1, preventing the throttle portion a of the first set, B, C, and the flow area of the closure start initial is rapidly reduced in D it can. That is, as shown by a two-dot chain line in FIG. 9, the constricted portions A, B, C, and D belonging to the first set are configured only by the flat surface S2 far from the grooves 51a and 51b of the second valve member 32. In this case, the distance L1 between the flat surface S2 at the beginning of closing of the throttle portions A, B, C, and D and the edges of the grooves 50a and 50b of the first valve member 31 is reduced. As a result, the steering assist force generating hydraulic pressure suddenly rises at the beginning of steering at low and medium speeds, and the rate of change of the steering assist force generating hydraulic pressure with respect to the steering torque changes suddenly as indicated by broken lines α ′ and γ ′ in FIG. Therefore, the steering feeling is lowered. Further, as shown by a broken line γ ′ during medium speed traveling, the steering assist force generating hydraulic pressure increases in the practical range of the steering assist force generating hydraulic pressure (the hydraulic pressure Pa or less in FIG. 9) even if the steering torque is small. Running stability is reduced. Furthermore, as indicated by a broken line β ′ in FIG. 8, the steering torque necessary to sufficiently increase the hydraulic pressure for generating the steering assist force during high speed traveling becomes excessive, and the necessary steering assist cannot be performed. On the other hand, according to the above configuration, the flat surface S1 at the beginning of closing of the throttle portions A, B, C, and D belonging to the first group and the edges of the grooves 50a, 50b, and 50c of the first valve member 31 9 can be increased by L2 in FIG. Therefore, the steering assist force generating hydraulic pressure does not increase rapidly at the beginning of steering during low and medium speed traveling, and the steering feeling can be improved by increasing the steering assist force generating hydraulic pressure at a substantially constant rate with respect to the steering torque. . In addition, during running at medium speed, it is possible to prevent the steering assist force generation hydraulic pressure from becoming excessive and improve running stability. Further, the steering assist necessary for sufficiently increasing the hydraulic pressure for generating the steering assist force during high speed traveling does not become excessive, and the necessary steering assist can be performed.
[0039]
FIG. 10 shows a first modification of the stepping motor 80. In the above embodiment, the rotation of the motor 80 is transmitted to the screw member 64 without decelerating, but in the first modification, the motor 80 is decelerated by the traction drive type decelerating mechanism 101. That is, the output shaft 80a of the motor 80 is inserted into the screw member 64 via the bush 109 so as to be relatively rotatable. The speed reduction mechanism 101 includes a support shaft 102 that is integrated with the screw member 64, a planetary roller 103 that is rotatably fitted to the support shaft 102, and a ring 104 that is integrated with the second valve housing 7 ′. Have The planetary roller 103 contacts the outer periphery of the output shaft 80a and the inner periphery of the ring 104. Thereby, the rotation of the output shaft 80a is transmitted to the planetary roller 103 by friction, and the screw member 64 is rotationally driven by the planetary roller 103 orbiting around the output shaft 80a by friction with the ring 104. . As a result, the output shaft 80a and the screw member 64 remain coaxially arranged without increasing the size of the motor 80, and the gear is used in a compact configuration compared to a reduction gear mechanism using a spur gear. The effect similar to that of increasing the output torque of the motor 80 can be obtained without any rotation transmission loss due to the backlash, and the operation at high and low temperatures can be performed reliably. Others are the same as that of the said embodiment.
[0040]
FIG. 11 shows a second modification of the stepping motor 80. In the above embodiment, the rotation of the motor 80 is transmitted to the screw member 64 without decelerating, but in the second modification, the planetary gear mechanism 201 decelerates. That is, the output shaft 80a of the motor 80 is inserted into the screw member 64 via the bush 209 so as to be relatively rotatable. The planetary gear mechanism 201 includes a support shaft 202 integrated with the screw member 64, a planetary gear 203 rotatably fitted on the support shaft 202, and a ring gear integrated with the second valve housing 7 ′. 204 and a sun gear 205 integrated with the output shaft 80a. The planetary gear 203 meshes with the ring gear 204 and the sun gear 205. As a result, the rotation of the output shaft 80a is transmitted from the sun gear 205 to the planetary gear 203, and the planetary gear 203 rotates around the output shaft 80a by meshing with the ring gear 204, so that the screw member 64 is rotationally driven. . As a result, the output torque of the motor 80 can be reduced with a compact configuration compared to a speed reduction mechanism using a spur gear while the output shaft 80a and the screw member 64 are arranged coaxially without increasing the size of the motor 80. The same effect as the increase can be obtained, and the operation at high and low temperatures can be performed reliably. Others are the same as that of the said embodiment.
[0041]
The present invention is not limited to the above embodiment. For example, in the above embodiment, the chamfered portion in the aperture portion belonging to the first group is configured from two flat surfaces arranged in parallel in the circumferential direction, but may be configured from three or more flat surfaces. Moreover, in the said embodiment, although the closing angle of the aperture | diaphragm | squeeze part which belongs to a 2nd group shall be mutually different, it may be the same. Further, the number of grooves in each valve member is not limited.
[0042]
【The invention's effect】
According to the present invention, it is possible to provide a hydraulic power steering apparatus that can improve steering feeling and running stability and can perform necessary steering assistance even during high-speed running.
[Brief description of the drawings]
FIG. 1 is a longitudinal sectional view of a hydraulic power steering apparatus according to an embodiment of the present invention. FIG. 2 is a diagram showing a hydraulic circuit of the hydraulic power steering apparatus according to the embodiment of the present invention. Explanatory drawing of the cross-sectional structure of the control valve in a steering apparatus. FIG. 4 is a development view of a control valve of a hydraulic power steering apparatus according to an embodiment of the present invention. FIG. 6 is a longitudinal sectional view of a variable throttle valve of the hydraulic power steering apparatus according to the embodiment of the present invention. FIG. 7 is a view illustrating the opening degree and the valve of the throttle portion of the control valve in the hydraulic power steering apparatus according to the embodiment of the present invention. FIG. 8 is a diagram showing a relationship between relative rotation angles of members. FIG. 8 is a diagram showing a relationship between steering torque and hydraulic pressure in a hydraulic power steering device. FIG. 9 is a hydraulic power steering according to an embodiment of the present invention. [Description of symbols] motor part cross-sectional view of a second modification of the first partial cross-sectional view of the motor of the modification [11] The present invention enlarged view FIG. 10 the present invention of a main part of the control valve of the ring device
20 Hydraulic cylinder 30 Control valve 31 First valve member 32 Second valve member 50a, 50b, 50c, 51a, 51b, 51c Groove 67 Variable restrictor 70 Pump 71 Tanks A, B, C, D Restrictors belonging to the first group Part A ′, B ′, C ′, D ′ The diaphragms S1, S2 belonging to the second set Flat surface

Claims

A hydraulic actuator (20) for generating a steering assist force, and a hydraulic control valve (30) acting on the actuator (20) ;
The control valve (30) includes a cylindrical first valve member (31) and a second valve member (32) inserted into the first valve member (31) so as to be relatively rotatable in accordance with a steering resistance. Have
A plurality of axial grooves (50a, 50b, 50c) are formed on the inner periphery of the first valve member (31) at intervals in the circumferential direction, and a plurality of grooves are formed on the outer periphery of the second valve member (32). Grooves (51a, 51b, 51c) along the axial direction are formed at intervals in the circumferential direction,
Between the edges along the axial direction of the grooves (50a, 50b, 50c) of the first valve member (31), a second groove in the valve member (32) (51a, 51b, 51c) and the edge along the axial direction of the Is a plurality of throttle portions (A, B, C, D, A ′, B ′, C ′, D ′) whose opening degree changes according to the relative rotation angle of both valve members (31, 32) ,
Via the control valve (30) , the steering assisting force according to the opening change of each throttle part (A, B, C, D, A ', B', C ', D') can be applied. An actuator (20) is connected to the pump (70) and the tank (71) ;
The edge along the axial direction of each groove (51a, 51b, 51c ) in the second valve member (32) is a chamfered portion, and the circumferential width of the chamfered portion is the respective narrowed portion (A, B, C, D). , A ′, B ′, C ′, D ′) are determined according to the closing angle, which is the relative rotation angle of both valve members (31, 32) required to fully close
Each aperture (A, B, C, D, A ′, B ′, C ′, D ′) is divided into a first set and a second set, and the apertures belonging to the second set (A ′, B ′, C ′, D ′) are made larger than the closing angles of the throttle parts (A, B, C, D) belonging to the first set,
Variable throttle part (67) whose opening degree changes in the oil passage between the throttle part (A ', B', C ', D') belonging to the second group and the tank (71) according to the operating conditions In the hydraulic power steering device provided with
The chamfered portion in the aperture portions (A, B, C, D) belonging to the first set is composed of a plurality of flat surfaces (S1, S2) arranged in parallel in the circumferential direction,
With respect to the radial direction of the circumscribed circle of the second valves members (32), the perpendicular direction of the tilt angle to each flat surface (S1, S2) are assumed to be different from each other,
The direction perpendicular to the flat surface (S1) closer to the groove (51a, 51b, 51c) of the second valve member (32) is the flat surface (S2) far from the groove (51a, 51b, 51c ). A hydraulic power steering apparatus, wherein the inclination angle is larger than a direction perpendicular to the angle.