JP3810232B2

JP3810232B2 - Power steering apparatus and servo valve used therefor

Info

Publication number: JP3810232B2
Application number: JP09704899A
Authority: JP
Inventors: 修佐野; 孝一夏成
Original assignee: JTEKT Corp
Current assignee: JTEKT Corp
Priority date: 1999-04-02
Filing date: 1999-04-02
Publication date: 2006-08-16
Anticipated expiration: 2019-04-02
Also published as: JP2000289633A

Description

【０００１】
【発明の属する技術分野】
本発明は、操舵に基づいて油圧ポンプからの圧油を操舵補助用の油圧シリンダの室へ切換えて供給するサーボ弁を備えた動力舵取装置及びこれに用いるサーボ弁に関する。
【０００２】
【従来の技術】
近年の自動車は、動力舵取装置、自動変速装置等、油圧により作動する多くの装置を備えており、これらの油圧作動装置に作動油を供給するための油圧ポンプが搭載されている。この種の油圧ポンプの多くは、エンジンの出力の一部を利用して駆動され、例えば、該エンジンの出力端に取り付けた動力取り出し用のプーリを介してのベルト伝動によりロータを回転駆動し、該ロータの回転に応じた油圧を発生する構成となっているが、このような単純な伝動構成を採用した場合、エンジンの回転速度が高い高速走行時に、油圧ポンプの駆動のための動力損失が大きく、燃費の低下を招き、また、この動力損失を軽減すべく油圧ポンプの仕様を決定した場合、エンジンの回転速度が低い低速走行時又は停車時に油圧ポンプの能力が不足し、油圧作動装置に十分な量の作動油を送給し得ないという問題がある。
【０００３】
このように自動車用の油圧ポンプにおいては、必要時に十分な作動油量を確保しつつ、駆動源となるエンジンの無為な動力損失を軽減することが要求されており、この要求に応えるべく本願出願人は、エンジンからロータへの伝動系の中途に、高低２段に切換え可能な変速手段と、前記ロータの回転に応じて吐出油路内に発生する油圧の作用により前記変速手段に切換え動作を行なわせる切換クラッチとを備え、舵輪の操舵によって操舵補助用の油圧シリンダが動作しているときには、これに伴って上昇する送油路内の油圧の作用により前記変速手段を高速側に切換え、操舵されることなく油圧シリンダが非動作のときには、前記変速手段を低速側に切換えるように構成した油圧ポンプユニットを特願平10-10770号に提案した。
【０００４】
ところがこのように構成された油圧ポンプユニットが組込された動力舵取装置にあっては、切換クラッチの動作により変速手段が低速側から高速側へ切換えられたとき、これに応じて油圧ポンプの回転速度が急変し、送油路内の油圧が急上昇することになり、この結果、操舵に基づいて前記油圧ポンプからの圧油を切換えて前記油圧シリンダへ供給するサーボ弁から油圧シリンダへの供給流量が急変し、油圧シリンダが発生する操舵補助力が急変し、舵輪の操作感覚が急変して、舵輪を把持するドライバに違和感を与えるという不具合がある。
【０００５】
このような問題を解消するため、本願出願人は、特願平10-10770号に提案したものを改良し、油圧ポンプに連通する前記送油路の中間をドレン油路に分岐する分岐部と、前記油圧シリンダへの供給流量に応じた差圧を発生する絞り手段と、前記分岐部に配してあり、前記差圧の作用により前記ドレン油路を開閉して、前記供給流量を制御する流量制御弁とを備えた油圧ポンプユニットを特願平10-126470 号に提案した。
【０００６】
この特願平10-126470 号にあっては、操舵されることなく油圧シリンダが非動作のときには、流量制御弁がドレン油路を開放し、油圧ポンプから送られる圧油の一部を分岐部を経てドレン油路に排出させた状態にあり、操舵によって油圧シリンダが動作し、送油路の内部の油圧が上昇し、切換クラッチの動作により変速手段が高速側に切換えられたとき、これに伴って生じる絞り手段の前後の差圧の増大に応じて流量制御弁が動作し、ドレン油路が閉止し、該ドレン油路への排油流量を徐々に減じられて油圧シリンダへの供給流量が漸増し、この供給油圧の急変を抑制することができるため、前記違和感を少なくすることができる。
【０００７】
【発明が解決しようとする課題】
ところで、以上の如く油圧ポンプの回転速度が高速と低速とに切換られるものにあっては、操舵されていないとき、油圧ポンプは低速で回転し、該油圧ポンプから油圧シリンダへ送られる供給流量よりもドレン油路への排油流量の方が多いため、入力トルク−圧力特性線が裾広がりの曲線を描くことになる。
【０００８】
従って、本願出願人が先に提案したように、操舵に伴って上昇する送油路内の油圧の作用により切換クラッチを動作させるように構成したものにあっては、操舵中立点から操舵され、送油路内の圧油が切換クラッチを動作させるために必要な所定値に上昇するまでの操舵角が所定の角度以上に大きくなり、操舵中立点から操舵されて油圧シリンダが動作するまでの操舵角が大となるため、操舵初期において操舵が重くなり過ぎることが懸念される。
【０００９】
本発明は斯かる事情に鑑みてなされたものであり、油圧ポンプから流量制御弁を経て圧油が供給されるサーボ弁の排油孔及び送油孔に絞り部を設けることによって、小流量の状態で操舵中立点から操舵が始まったとき、排油孔及び該排油孔に連通する送油孔による排油流量を早期に少なくすることができるとともに、サーボ弁から油圧シリンダへの供給流量を早期に多くでき、その油圧力を早期に所定値に上昇させることができ、操舵初期において油圧シリンダが動作するまでの操舵角を小さくでき、操舵を軽くすることができる動力舵取装置、及びサーボ弁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】
第１発明に係る動力舵取装置は、油圧ポンプと、該油圧ポンプに送油路を介して連通する操舵補助用の油圧シリンダと、前記送油路中に設けられる流量制御弁と、操舵に基づいて前記流量制御弁からの圧油を前記油圧シリンダの室へ切換えて供給するサーボ弁とを備え、該サーボ弁には、バルブボディに相対角変位を可能として嵌合されるバルブスプール及び前記バルブボディの間に、給油孔に連通する給油室、第１中間圧室、第２中間圧室及び第１中間圧室がこの順序で複数個設けられている動力舵取装置において、前記第２中間圧室は、少なくとも１つが排油孔に連通し、他の少なくとも一つが絞り部を有する排油孔に連通されており、該絞り部を有する排油孔に連通する第２中間圧室の両側で隣り合う第１中間圧室は絞り部を有する送油孔に連通されており、該第１中間圧室を除く残りの第１中間圧室は送油孔に連通していることを特徴とする。
【００１１】
第２発明に係る動力舵取装置は、駆動源に変速手段を介して連動する油圧ポンプと、該油圧ポンプに送油路を介して連通する操舵補助用の油圧シリンダと、前記送油路中に設けられる流量制御弁と、操舵に基づいて前記流量制御弁からの圧油を前記油圧シリンダの室へ切換えて供給するサーボ弁と、操舵に基づいて前記変速手段に変速の動作を行わせる切換クラッチとを備え、前記サーボ弁には、バルブボディに相対角変位を可能として嵌合されるバルブスプール及び前記バルブボディの間に、給油孔に連通する給油室、第１中間圧室、第２中間圧室及び第１中間圧室がこの順序で複数個設けられている動力舵取装置において、前記第２中間圧室は、少なくとも１つが排油孔に連通し、他の少なくとも一つが絞り部を有する排油孔に連通されており、該絞り部を有する排油孔に連通する第２中間圧室の両側で隣り合う第１中間圧室は絞り部を有する送油孔に連通されており、該第１中間圧室を除く残りの第１中間圧室は送油孔に連通していることを特徴とする。
【００１２】
第３発明に係るサーボ弁は、バルブボディに相対角変位を可能として嵌合されるバルブスプール及び前記バルブボディの間に、給油孔に連通する給油室、第１中間圧室、第２中間圧室及び第１中間圧室がこの順序で複数個設けられているサーボ弁において、前記第２中間圧室は、少なくとも１つが排油孔に連通し、他の少なくとも一つが絞り部を有する排油孔に連通されており、該絞り部を有する排油孔に連通する第２中間圧室の両側で隣り合う第１中間圧室は絞り部を有する送油孔に連通されており、該第１中間圧室を除く残りの第１中間圧室は送油孔に連通していることを特徴とする。
【００１３】
第１発明、第２発明及び第３発明にあっては、操舵されていないとき、油圧シリンダに加わる負荷の大きさは小さいため、流量制御弁を経てサーボ弁へ供給された圧油は、その全流量が絞り部を有する排油孔及び有しない排油孔から油タンク等へ還流される。この小流量の状態で操舵中立点から操舵が始まったとき、油圧シリンダに加わる負荷の大きさが漸増し、流量制御弁からサーボ弁への供給流量が漸増するとともに、サーボ弁のバルブスプールが回転し、その弁開度が漸増してサーボ弁へ供給される流量が漸増する。
【００１４】
このとき、第２中間圧室の少なくとも１つが排油孔に連通し、他の少なくとも一つが絞り部を有する排油孔に連通しているとともに、絞り部を有する排油孔に連通する第２中間圧室の両側で隣り合う第１中間圧室が絞り部を有する送油孔に連通しているため、全ての排油孔の開口面積及び送油孔の開口面積が同じとされた既存のサーボ弁を備えたものに比較して排油孔及び該排油孔に連通する送油孔による排油流量を早期に少なくすることができるとともに、サーボ弁から油圧シリンダへの供給流量を早期に多くでき、その油圧力を早期に所定値に上昇させることができる。従って、小流量の状態で操舵中立点から操舵されて油圧シリンダが動作を開始するまでの操舵角を小さくでき、操舵初期において操舵を軽くすることができる。
【００１５】
また、第１発明及び第２発明にあっては、サーボ弁から油圧シリンダへ供給する圧油の油圧力が所定値に上昇した後、流量制御弁の弁特性に従って流量が増加することになるのであり、従って、油圧シリンダが動作を開始するとき、油圧力の不連続部をなくすことができ、操舵を行なっている運転者に違和感を与えることをなくし得るのである。
【００１６】
【発明の実施の形態】
以下本発明をその実施例を示す図面に基づいて詳述する。
実施の形態１
図１は動力舵取装置の油圧系統図、図２は油圧ポンプの回転速度を切換える切換え部分の断面図である。
【００１７】
本発明の動力舵取装置は、自動車のエンジンである駆動源１に変速手段２を介して連動する油圧ポンプ３と、該油圧ポンプ３の吐出部に送油路４を介して連通する操舵補助用の油圧シリンダ５と、前記送油路４の中間に設けられ、前記油圧シリンダ５に加わる負荷の大きさに応じて前記油圧シリンダ５へ送る圧油の流量を制御する流量制御弁６と、操舵に基づいて前記流量制御弁６からの圧油を前記油圧シリンダ５の室へ切換えて供給するサーボ弁７と、操舵に基づいて前記変速手段２に変速の動作を行わせる切換クラッチ８とを備えている。
【００１８】
変速手段２は、例えば遊星歯車変速機、遊星ローラ変速機等、油圧ポンプ３の回転速度を例えば高速と低速とに変速するものである。
【００１９】
流量制御弁６は、送油路４の中途に、これと交叉する態様に円形断面を有するスプール室６０が形成されており、該スプール室６０の内部に流量制御弁６が構成されている。
【００２０】
図３は流量制御弁６部分の概略構成を示す断面図である。スプール室６０は、図２、図３に示す如く、切換クラッチ８の圧油室９への導圧のために設けられた導油路１０よりも上流側において、その長手方向の一側を送油路４の中途に連通させて形成してある。またスプール室６０の他側は、これと分岐部との間において、送油路４の中間に設けられた絞り手段、詳しくは固定絞り６１の下流側に導圧路６２を介して連通させてあり、更にスプール室６０の中間部は、該位置に開口するドレン油路６３を介して低圧に保たれた油圧ポンプ３の吸込油路１１（図２参照）に連通させてある。
【００２１】
流量制御弁６は、以上の如きスプール室６０の内部に、軸長方向への摺動可能に嵌挿された有底円筒形をなす第１のスプール６ａと、該第１のスプール６ａの内側に軸長方向への摺動自在に嵌挿された円柱形をなす第２のスプール６ｂとを備えてなる。第１のスプール６ａの底部には、所定の内径を有する貫通孔が形成されており、この貫通孔には、第２のスプール６ｂの一側に同軸的に突設された連結ロッド６ｃが内嵌され、第１のスプール６ａの外側に突出する連結ロッド６ｃの端部には、第２のスプール６ｂと等しい外径を有する受圧板６ｄが同軸的に固定されている。
【００２２】
第１のスプール６ａと第２のスプール６ｂとは、前者の内底面と、これに対向する後者の端面との間に介装された付勢ばね６４により互いに離反する方向に付勢して一体化され、第１のスプール６ａの開口側を送油路４との連通側に向け、同じく底面側を前記導圧路６２との連通側に向けてスプール室６０の内部に装着されており、前記受圧板６ｄと、これに対向するスプール室６０の端面との間に介装された制御ばね６５により、前記端面から離反する方向、即ち、送油路４との連通側に向けて付勢され、第１のスプール６ａの開口側周縁部が前記ドレン油路６３を塞ぐように位置決めされている。
【００２３】
即ち、スプール室６０の内部は、第１のスプール６ａの開口側に面し、前記送油路４に連通された第１圧力室６０ａと、第１のスプール６ａの底部側に面し、前記導圧路６２が連通された第２圧力室６０ｂとに分割されており、前記ドレン油路６３は、第１のスプール６ａの開口側周縁部により、第１圧力室６０ａとの連通を遮断された状態にある。
【００２４】
また第１のスプール６ａ内側の前記付勢ばね６４の配設室は、該第１のスプール６ａの周面を貫通する貫通孔６６によりドレン油路６３の開口部に連通されており、第１のスプール６ａに対して後述の如く生じる第２のスプール６ｂの摺動が、前記貫通孔６６を経て生じる油の出入りにより支障なく行なわれるようになしてある。
【００２５】
図４及び図５は以上の如く構成された流量制御弁６の動作説明図である。油圧ポンプ３の動作により吐出される圧油は、これらの図中に矢符により示す如く、送油路４を経てスプール室６０の一側の第１圧力室６０ａに流入し、スプール室６０の下流側の送油路４の中間に配した前記固定絞り６１を経て油圧シリンダ５に送給される。
【００２６】
このとき前記第１圧力室６０ａの内部には、油圧Ｐ1 の圧油が導入され、スプール室６０の他側の第２圧力室６０ｂには、前記導圧路６２を介して固定絞り６１の下流側の油圧Ｐ2 が導入される。この油圧Ｐ2 は、油圧Ｐ1 を有して送油路４に送り出される圧油が固定絞り６１の通過により減圧されて生じ、この減圧量は、固定絞り６１を通過する圧油量、即ち、油圧シリンダ５への供給流量の増加に応じて大となるから、第１圧力室６０ａと第２圧力室６０ｂとの間には、前記供給流量の増加に伴って増大する差圧ΔＰ（＝Ｐ1 −Ｐ2 ）が発生する。
【００２７】
ここで第２のスプール６ｂは、第１圧力室６０ａの油圧Ｐ1 を第１のスプール６ａの開口側に露出する一側端面に受圧し、また第２圧力室６０ｂの油圧Ｐ2 を同側に一体化された前記受圧板６ｄの端面に受圧するが、この受圧板６ｄは、前述の如く、第２のスプール６ｂと略等しい外径を有しており、第１圧力室６０ａ及び第２圧力室６０ｂ側の受圧面積は相等しい。従って第２のスプール６ｂは、前記差圧ΔＰの増大に伴って第１圧力室６０ａから第２圧力室６０ｂに向けて押圧され、この押圧力と前記制御ばね６５のばね力との力バランスに応じて第２圧力室６０ｂの側に向けて移動する。
【００２８】
これに対し第１のスプール６ａは、第１圧力室６０ａの油圧Ｐ1 を開口側の周縁に受圧し、また第２圧力室６０ｂの油圧Ｐ2 を底面に受圧することとなり、これらの受圧面積は後者の方が大きい。従って第１のスプール６ａは、前記差圧ΔＰの作用ではなく、底面側に加わる第２圧力室６０ｂの油圧Ｐ2 、即ち、固定絞り６１の下流側の油圧Ｐ2 の作用により第２圧力室６０ｂから第１圧力室６０ａに向けて押圧され、この押圧力と前記付勢ばね６４のばね力との力バランスに応じて、前記第２のスプール６ｂに対して第１圧力室６０ａの側に向けて相対移動する。
【００２９】
操舵されることなく油圧シリンダ５が非動作のとき等、固定絞り６１の下流側に発生する油圧Ｐ2 が十分に小さい場合、第１圧力室６０ａの油圧Ｐ1 、第２圧力室６０ｂの油圧Ｐ2 、及びこれらの間の差圧ΔＰのいずれもが小さいことから、第１のスプール６ａ及び第２のスプール６ｂは、図３に示す組み立て位置から左方に一体をなして移動し、ドレン油路６３がわずかに開放された状態となり、油圧ポンプ３から送油路４内に送出される圧油は、その大部分が固定絞り６１を経て油圧シリンダ５に送給される。
【００３０】
この状態から操舵によって油圧シリンダ５が動作し、油圧Ｐ2 が上昇した場合、固定絞り６１の前後の差圧ΔＰもまた増大する結果、第２のスプール６ｂは、前記差圧ΔＰの作用による押圧力が制御ばね６５のばね力を上回ると共に、このばね力に抗して外側の第１のスプール６ａと一体をなして移動を開始し、図４に示す如く、この移動に応じてスプール室６０の内部へのドレン油路６３の開口面積が大きくなり、油圧ポンプ３から送油路４内に供給される供給流量の大部分がドレン油路６３を経て低圧側に排油せしめられるようになる。
【００３１】
このときの排油流量は、固定絞り６１の前後の差圧ΔＰの増加、即ち、供給流量の増加に伴って増すから、油圧シリンダ５への供給流量は、油圧ポンプ３の回転速度の上昇に拘わらず略一定に保たれる。
【００３２】
一方、前述した如く、第２圧力室６０ｂの油圧Ｐ2 の作用により押圧されている第１のスプール６ａは、該油圧Ｐ2 による押圧力が付勢ばね６４のばね力を上回ると共に第２のスプール６ｂに対する前述した相対移動を開始する。これにより図５に示す如く、スプール室６０の内部へのドレン油路６３の開口面積が減少せしめられ、該ドレン油路６３を経て低圧側に排油される排油流量が減少し、油圧シリンダ５への供給流量は、油圧Ｐ2 の上昇に応じて増加するようになる。
【００３３】
図６は、流量制御弁６の下流側圧油の圧力とその流量との関係を示す流量特性図である。図示の如く油圧ポンプ３からの供給油量は、下流側の油圧Ｐ2 が小さく、流量制御弁６において図３〜図４に示す状態変化が生じている間にはイの如く略一定に保たれ、前記油圧Ｐ2 が所定の限界圧力Ｐs を超え、流量制御弁６において図４〜図５に示す状態変化が生じている間は、前記油圧Ｐ2 の上昇に伴ってロの如く増加し、前記油圧Ｐ2 が所定の油圧力Ｐに上昇したとき前記下流側の流量がハの如く大流量に復帰する。従って、前記限界圧力ＰS を、切換クラッチ８の動作により変速手段２が高速側に切換えられる回転速度と略一致するように設定することにより、この切換えの前後に生じる油圧の急変を防止することができる。なお、限界圧力ＰS の前述した設定は、第１のスプール６ａの両側の受圧面積差、及び付勢ばね６４のばね力を適正に選定することにより実現できる。
【００３４】
サーボ弁７は、入力軸を介して舵輪に繋がるバルブスプール７ａと、該バルブスプール７ａの周りに挿嵌され、出力軸に繋がるバルブボディ７ｂとを備える。これらバルブスプール７ａ及びバルブボディ７ｂの基本的な構成は既存のものと同じであり、実施の形態１では８等配のサーボ弁を用いている。円筒形をなすバルブボディ７ｂの内周面には、夫々等しい幅を有する８個の第１の油溝７０…が周方向に等配をなして並設され、また、バルブボディ７ｂの内径と略等しい外径を有する厚肉円筒形のバルブスプール７ａの外周面には、同様に、夫々等しい幅を有する８個の第２の油溝７１…が周方向に等配をなして並設されている。
【００３５】
バルブスプール７ａはバルブボディ７ｂの内側に同軸上での相対回転を可能として嵌合し、これら両者を、バルブスプール７ａの内側に挿通されたトーションバー１２により相互に連結して構成されている。第１の油溝７０…と第２の油溝７１…とは、前記トーションバー１２に捩れが生じていない中立状態において、図示の如く周方向に千鳥配置され、夫々の両側に相隣するものと連通するように位置決めされている。
【００３６】
以上の構成により、バルブボディ７ｂの第１の油溝７０…の夫々は、バルブスプール７ａの第２の油溝７１…間のランドに対向し、また、バルブスプール７ａの第２の油溝７１…の夫々は、バルブボディ７ｂの第１の油溝７０…間のランドに対向して、バルブボディ７ｂとバルブスプール７ａとの嵌合周上には、第１の油溝７０…の内側（幅方向両側の溝縁間）の８つの油室と、第２の油溝７１…の外側（幅方向両側の溝縁間）の８つの油室とが、夫々の間に連通部を有して交互に並んだ状態となる。
【００３７】
バルブボディ７ｂとバルブスプール７ａとは、これらを連結するトーションバー１２の捩れの範囲内での相対角変位が可能であり、前記各油室間の連通部、即ち、油溝７０…，７１…の幅方向両側の溝縁間は、前記相対角変位に応じて夫々の連通面積（絞り面積）を増減する絞り部ａ，ｂとして作用する。
【００３８】
バルブスプール７ａの第２の油溝７１…により形成された８つの油室の内、１つおきに位置する４つは、バルブボディ７ｂの周壁を貫通し、夫々の油溝７１…の外側に開口を有する各別の給油孔７２…を介して前記流量制御弁６の下流側に前記固定絞り６１を介して接続され、油圧ポンプ３から圧油の供給がなされる給油室７３…を構成している。これに対し、残りの４つの油室は、バルブスプール７ａを半径方向に貫通し、夫々の油溝７１…の底部に開口を有する各別の排油孔７４…及びバルブスプール７ａ内側の中空部を介して排油先となる油タンク１３に接続され、該油タンク１３への排出油の通路となる第２中間圧室（以下排油室という）７５…を構成している。
【００３９】
一方、第１の油溝７０…の内側に形成された８つの油室の内、前記給油室７３…に周方向の同側にて相隣する４つの油室は、バルブボディ７ｂの周壁を貫通し、夫々の油溝７０…の底部に開口を有する各別の送油孔７６…を介して圧油の送給先である油圧シリンダ５の一方のシリンダ室５R に接続され、このシリンダ室５R への第１中間圧室（以下第１の送油室という）７７…を構成しており、残りの４つは、同様の送油孔７８…を介して前記油圧シリンダ５の他方のシリンダ室５L に接続され、該シリンダ室５L への第１中間圧室（以下第２の送油室という）７９…を構成している。従って、給油室７３…の両側には、第１の送油室７７又は第２の送油室７９を経て排油室７５に至る油路が夫々形成され、給油室７３と送油室７７及び排油室７５と送油室７９とが絞り部ａを介して連通し、また、給油室７３と送油室７９及び排油室７５と送油室７７とが絞り部ｂを介して連通される。
【００４０】
以上の如く構成されたサーボ弁７は既存の８等配タイプの基本的な構成であり、本発明は前記した如く偶数個の排油孔７４の１つ置きに絞り部８０を設け、各給油孔７２…に対し一側に絞り部８０を有する第１の排油孔７４ａが、他側に絞り部８０を有しない第２の排油孔７４ｂが夫々配設された構成とし、さらに、絞り部８０を有する第１の排油孔７４ａに連通する排油室７５の両側で隣り合う第１及び第２の送油室７７，７９に連通する送油孔７６，７８に絞り部８１，８１を設けてこれを第１の送油孔７６ａ，７８ａとし、絞り部８０を有しない第２の排油孔７４ｂに連通する排油室７５の両側で隣り合う第１及び第２の送油室７７，７９に連通する送油孔７６，７８に絞り部を設けなくてこれを第２の送油孔７６ｂ，７８ｂとし、全ての給油孔７２…へ供給された圧油の排油流量が、第１及び第２の排油孔７４ａ，７４ｂによる排油流量及びこれら排油孔７４ａ，７４ｂに連通する送油孔７６ａ，７６ｂ又は送油孔７８ａ，７８ｂを平均した平均排油流量となるようにした。
【００４１】
絞り部８０，８１は、パイプ等の筒部材を用いてなり、絞り部８０を既存のサーボ弁７の排油孔７４に挿嵌し、該絞り部８０を設けた第１の排油孔７４ａの排油流量を、絞り部８０を設けない第２の排油孔７４ｂの排油流量に比較して少なくする。また、絞り部８１を既存のサーボ弁７の送油孔７６又は送油孔７８に挿嵌し、該絞り部８１を設けた第１の送油孔７６ａ，７８ａの排油流量を、絞り部８１を設けない第２の送油孔７６ｂ，７８ｂの排油流量に比較して少なくする。
【００４２】
図７は操舵角とサーボ弁７の排油孔及び送油孔の開口面積との関係を示す油圧特性図であって、(a) の破線は本発明の構成であり、(b) 及び(a) の実線イは既存の構成であり、(c) 及び(a) の実線ロは既存の構成の排油孔の全てに絞り部８０を設けるとともに、送油孔の全てに絞り部８１を設けた構成である。全ての排油孔７４及び全ての送油孔７６，７８に絞り部８０，８１を有しない場合、図７(b) 及び(a) の実線イの如く操舵初期時における開口面積の減少量は比較的多いのであり、また、全ての排油孔７４及び送油孔７６，７８に絞り部８０，８１が設けられた場合、図７(c) 及び(a) の実線ロの如く操舵初期時における開口面積の減少量は一定であり、何ら変わらないのに対し、少なくとも１つの排油孔７４に絞り部８０を設けて第１及び第２の排油孔７４ａ，７４ｂとし、さらに、絞り部８０を有する第１の排油孔７４ａに連通する排油室７５の両側で隣り合う第１及び第２の送油室７７，７９に連通する送油孔７６，７８に絞り部８１，８１を設けて第１の送油孔７６ａ，７８ａ、第２の送油孔７６ｂ，７８ｂとした本発明にあっては、図７(a) の破線の如く前記(b) 及び(c) を平均化した開口面積に変えることができる。
【００４３】
図８は操舵角とサーボ弁によって制御される油圧力との関係を示す油圧特性図である。
このように全ての給油孔７２に対する排油孔７４ａ，７４ｂ及び送油孔７６ａ，７６ｂ又は送油孔７８ａ，７８ｂを平均化した開口面積に変えることによって、流量制御弁６によってその下流側の流量が所定の油圧力Ｐに上昇するために必要な操舵角をθとした場合、操舵角がθになったときの排油孔７４ａ，７４ｂ及び送油孔７６ａ，７６ｂ又は送油孔７８ａ，７８ｂの開口面積を前記絞り部８０，８１の開口面積に近付けることによって、操舵角がθ度未満では絞り部８０，８１の影響を受けて排油され、小流量に対応した油圧特性が得られ、操舵角がθ度以上では絞り部８０，８１の影響を受けずに排油され、既存の８等配サーボ弁の油圧特性、すなわち図６のハに示すような大流量が得られる。
【００４４】
図９、図１０は操舵が始まったときのサーボ弁７の動作説明図である。
例えば、図９に示す如く操舵中立点から右方向へ操舵されたとき、バルブスプール７ａがバルブボディ７ｂに対し時計方向へ回転し、絞り部ａが漸次開放されるとともに、絞り部ｂが漸次閉じられ、全ての給油孔７２の両側に配置されている排油孔７４ａ，７４ｂから直接排油されるとともに、全ての給油孔７２に対し反時計方向側に配置されている送油孔７８ａ，７８ｂを介して前記排油孔７４ａ，７４ｂから排油されながら送油孔７６ａ，７６ｂへ供給されることになるため、前記絞り部ａ，ｂによる絞り量を変えることによって操舵角がθになったときの排油孔７４ａ，７４ｂ及び送油孔７８ａ，７８ｂの開口面積を前記絞り部８０，８１の開口面積に近付けることができる。
【００４５】
また、図１０に示す如く操舵中立点から左方向へ操舵されたとき、バルブスプール７ａがバルブボディ７ｂに対し反時計方向へ回転し、絞り部ｂが漸次開放されるとともに、絞り部ａが漸次閉じられ、全ての給油孔７２の両側に配置されている排油孔７４ａ，７４ｂから直接排油されるとともに、全ての給油孔７２に対し時計方向側に配置されている送油孔７６ａ，７６ｂを介して前記排油孔７４ａ，７４ｂから排油されながら送油孔７８ａ，７８ｂへ供給されることになるため、前記絞り部ａ，ｂによる絞り量を変えることによって操舵角がθになったときの排油孔７４ａ，７４ｂ及び送油孔７６ａ，７６ｂの開口面積を前記絞り部８０，８１の開口面積に近付けることができる。
【００４６】
次に以上の如く構成された動力舵取装置の動作について説明する。
アイドリング時など舵輪が操舵されていないとき、油圧シリンダ５に加わる負荷の大きさは小さいため、切換クラッチ８によって変速手段２が低速回転側へ切換えられ、油圧ポンプ３は低速回転であり、該油圧ポンプ３から送油路４へ供給される圧油の流量は図６イの如く少量である。従って、流量制御弁６を経てサーボ弁７へ供給された圧油は、その全流量が絞り部８０を有する第１の排油孔７４ａ及び絞り部８０を有しない第２の排油孔７４ｂから油タンク１３等へ還流される。
【００４７】
給油孔７２から給油室７３に導油される流量は小流量であり、この小流量の圧油が４等配されることになり、４個所の給油室７３に分配された圧油は、４個所の給油室７３の両側の油路に均等に配分され、第１の送油室７７及び第２の送油室７９を経て排油室７５に達し、これら夫々に開口する第１及び第２の排油孔７４ａ，７４ｂを経てバルブスプール７ａ内側の中空部に流れ込み、該中空部内にて合流して油タンク１３に排油される。従って、前記送油室７７，７９間及びこれら夫々に接続された油圧シリンダ５の両シリンダ室ＳR ，ＳL 間に圧力差は発生せず、該油圧シリンダ５はなんらの力も発生しない。
【００４８】
この小流量の状態で操舵中立点から操舵されたとき、油圧シリンダ５に加わる負荷の大きさが漸増することになり、これに伴い流量制御弁６によってサーボ弁７へ供給される圧油の流量が漸増することになるとともに、図９又は図１０の如くサーボ弁７のバルブスプール７ａが回転し、その弁開度が漸増して給油孔７２へ供給される流量が漸増する。このとき、少なくとも一つの排油孔７４には絞り部８０が、また、この絞り部８０を有する排油孔７４ａの両側に配置される１組の送油孔７６，７８には絞り部８１が夫々設けられているため、全ての排油孔７４の開口面積及び全ての送油孔７６，７８の開口面積が同じとされた既存のサーボ弁を備えたものに比較して排油孔７４による排油流量を早期に少なくすることができるとともに、送油孔７６又は送油孔７８への供給流量を早期に多くでき、その油圧力を油圧シリンダ５を動作させるのに必要な所定値Ｐｓに早期に上昇させることができる。
【００４９】
従って、小流量の状態で操舵中立点から操舵されて油圧シリンダ５が動作を開始するまでの操舵角を小さくでき、操舵初期において操舵を軽くすることができる。
【００５０】
このように送油孔７６又は送油孔７８内の油圧力が所定値ＰＳに上昇したとき、切換クラッチ８が動作し、変速手段２が高速側に切換えられると共に流量制御弁６の前述した動作が行なわれるから、油圧シリンダ５へ供給される圧油の油圧力は、図６ロの如く前記切換え前の回転速度での変化状態から、切換え後の回転速度での変化状態に連続して漸近する特性を示し、不連続部をなくすことができる。また、小流量の状態で操舵中立点から操舵されて油圧シリンダ５が動作を開始するまでの操舵角を小さくできるため、少流量曲線Ａから多流量曲線Ｂへの移行長さＨを本発明出願人が先に提案した移行長さＨ１に比較して短くでき、油圧シリンダ５が発生する操舵補助力は、前記所定値Ｐｓの前後において急変することがなく、操舵を行なっている運転者に違和感を与えることがない。
【００５１】
実施の形態２（参考例）
図１１は動力舵取装置の油圧系統図である。
この実施の形態２の動力舵取装置及びこれに用いるサーボ弁７は参考例であり、前記送油孔７６，７８に絞り部８１を設ける代わりに、絞り部８０を有する排油孔７４ａに連通する排油室７５の両側で隣り合う１組の送油室７７，７９の送油孔７６，７８との連通を遮断、換言すれば実施の形態１の送油孔７６ａ，７８ａをなくし、全ての給油孔７２…へ供給された圧油の排油流量が、第１及び第２の排油孔７４ａ，７４ｂによる排油流量を平均した平均排油流量となるようにしたものであり、その他の構成及び作用は実施の形態１と同じであるため、共通部品については同じ符号を付し、その詳細な説明及び構造、作用、効果を省略する。
【００５２】
送油室７７，７９の送油孔７６，７８との連通を遮断するのは、例えば既存の前記バルブボディ７ｂに少なくとも１組の送油孔７６，７８を穿設しないか、又は、既存の前記バルブボディ７ｂに穿設された、絞り部８０を有する排油孔７４ａに連通する排油室７５の両側で隣り合う送油孔７６，７８に埋体を挿嵌して閉鎖するのである。
【００５３】
この実施の形態２にあっては、全ての給油孔７２…へ供給された圧油は、絞り部８０を有する排油孔７４ａに連通する排油室７５の両側で隣り合う１組の送油室７７，７９から送油孔７６，７８を経て排油孔７４ａ，７４ｂへ排油されず、全ての給油孔７２…へ供給された圧油は、直接排油孔７４ａ，７４ｂへ排油されることになり、実施の形態１と同様の作用効果が得られる。
【００５４】
実施の形態３（参考例）
図１２は動力舵取装置の油圧系統図である。
この実施の形態３の動力舵取装置及びこれに用いるサーボ弁７は参考例であり、前記排油孔７４に絞り部８０を設ける代わりに、少なくとも一つの前記排油室７５の排油孔７４との連通を遮断、換言すれば実施の形態１において絞り部８０を設けた前記排油孔７４ａをなくし、該排油孔７４との連通を遮断した排油室７５の両側で隣り合う１組の送油室７７，７９を、絞り部８１を有する前記送油孔７６ａ，７８ａに連通させ、全ての給油孔７２…へ供給された圧油の排油流量が、前記排油孔７４及び前記送油孔７６ａ，７６ｂ又は前記送油孔７８ａ，７８ｂによる排油流量を平均した平均排油流量となるようにしたものであり、その他の構成及び作用は実施の形態１と同じであるため、共通部品については同じ符号を付し、その詳細な説明及び構造、作用、効果を省略する。
【００５５】
排油室７５の排油孔７４との連通を遮断するのは、例えば既存の前記バルブスプール７ａに少なくとも一つの排油孔７４を穿設しないか、又は、既存の前記バルブスプール７ａに穿設された、絞り部８１を有する送油孔７６ａ，７８ａに連通する送油室７７，７９の間にある排油室の排油孔７４に埋体を挿嵌して閉鎖するのである。
【００５６】
この実施の形態３にあっては、全ての給油孔７２…へ供給された圧油は、排油孔７４から直接排油されるとともに、排油孔７４との連通が遮断された排油室７５の両側で隣り合う１組の送油室７７，７９から送油孔７６ａ，７６ｂ又は送油孔７８ａ，７８ｂを経て排油孔７４へ排油されることになり、実施の形態１と同様の作用効果が得られる。
【００５７】
実施の形態４
この実施の形態４の動力舵取装置及びこれに用いるサーボ弁７は、図示していないが、実施の形態１，２の排油孔７４にバルブスプール７ａと別個に形成された絞り部８０を挿嵌して設けるとともに、実施の形態１，３の送油孔７６，７８にバルブボディ７ｂと別個に形成された絞り部８１を挿嵌して設ける代わりに、絞り部８０がバルブスプール７ａと一体に形成され、さらに、絞り部８１がバルブボディ７ｂと一体に形成された構成としたものであり、その他の構成及び作用は実施の形態１，２，３と同じであるため、その詳細な説明及び構造、作用、効果を省略する。この実施の形態４にあっては、既存のサーボ弁を用いて安価に構成することができる。
【００５８】
なお、本発明に係る動力舵取装置は、変速手段２に代えて、例えば小流量タイプの油圧ポンプ及び大流量タイプの油圧ポンプ３を備えた構成とし、舵輪が操舵されていないときは小流量タイプの油圧ポンプを選定し、操舵中立点から操舵されて油圧シリンダ５を動作させるのに必要な所定値に上昇したとき、大流量タイプの油圧ポンプを選定する如く構成してもよいのであり、要は小流量と大流量とに変えることができればよい。
【００５９】
また、以上の実施の形態１〜４において、サーボ弁７は、複数個の排油孔７４を有する４等配以上、例えば６等配、１２等配、１４等配、１６等配等の偶数個の排油孔、好ましくは８等配以上が好ましいのであるが、その他、奇数個の排油孔を有するサーボ弁であってもよいのであり、要は排油孔の少なくとも１つが排油孔７４に連通し、他の少なくとも一つが排油孔７４との連通を遮断又は絞り部８０を有する排油孔７４ａに連通されており、該遮断された排油室（第２中間圧室）７５又は絞り部８１を有する排油孔に連通する排油室（第２中間圧室）７５の両側で隣り合う１組の送油室（第１中間圧室）７７，７９は送油孔７６，７８との連通を遮断又は絞り部８１を有する送油孔７６ａ，７８ａに連通されており、該送油室を除く残りの送油室は送油孔７６，７８に連通している構成であればよい。
【００６０】
【発明の効果】
以上詳述した如く第１発明、第２発明及び第３発明によれば、全ての排油孔の開口面積が同じであり、さらに、全ての送油孔の開口面積が同じとされた既存のサーボ弁を備えたものに比較して排油孔による排油流量を早期に少なくすることができるとともに、サーボ弁から油圧シリンダへの供給流量を早期に多くでき、その油圧力を早期に所定値に上昇させることができるため、小流量の状態で操舵中立点から操舵されて油圧シリンダが動作を開始するまでの操舵角を小さくでき、操舵初期において操舵を軽くすることができる。
【００６１】
しかも、第１発明及び第２発明によれば、油圧シリンダが動作を開始するとき、油圧力の不連続部をなくすことができ、舵輪操作を行なっている運転者に違和感を与えることをなくし得るのである。
【図面の簡単な説明】
【図１】本発明に係る動力舵取装置の実施の形態１の油圧系統図である。
【図２】本発明に係る動力舵取装置の油圧ポンプの回転速度を切換える切換え部分の断面図である。
【図３】本発明に係る動力舵取装置の流量制御弁部分の概略構成を示す断面図である。
【図４】本発明に係る動力舵取装置の流量制御弁の動作説明図である。
【図５】本発明に係る動力舵取装置の流量制御弁の動作説明図である。
【図６】本発明に係る動力舵取装置の流量制御弁の下流側圧油の圧力とその流量との関係を示す流量特性図である。
【図７】本発明に係る動力舵取装置の操舵角とサーボ弁の排油孔の開口面積との関係を示す油圧特性図である。
【図８】本発明に係る動力舵取装置の操舵角とサーボ弁によって制御される油圧力との関係を示す油圧特性図である。
【図９】本発明に係る動力舵取装置の操舵が始まったときのサーボ弁の動作説明図である。
【図１０】本発明に係る動力舵取装置の操舵が始まったときのサーボ弁の動作説明図である。
【図１１】参考例に係る動力舵取装置の実施の形態２の油圧系統図である。
【図１２】参考例に係る動力舵取装置の実施の形態３の油圧系統図である。
【符号の説明】
１駆動源
２変速手段
３油圧ポンプ
４送油路
５油圧シリンダ
６流量制御弁
７サーボ弁
７ａバルブスプール
７ｂバルブボディ
７２給油孔
７３給油室
７４排油孔
７５排油室
７６，７８送油孔
７７，７９送油室
８切換クラッチ
８０絞り部
８１絞り部[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a power steering apparatus having a servo valve that supplies pressure oil from a hydraulic pump to a chamber of a hydraulic cylinder for assisting steering based on steering, and a servo valve used therefor.
[0002]
[Prior art]
In recent years, automobiles are provided with many devices that operate by hydraulic pressure, such as a power steering device and an automatic transmission, and are equipped with a hydraulic pump for supplying hydraulic fluid to these hydraulic operation devices. Most of this type of hydraulic pump is driven by utilizing a part of the output of the engine, for example, the rotor is driven to rotate by belt transmission through a pulley for taking out power attached to the output end of the engine, The hydraulic pressure is generated according to the rotation of the rotor. However, when such a simple transmission configuration is adopted, the power loss for driving the hydraulic pump is reduced when the engine is running at a high speed. If the specifications of the hydraulic pump are determined to reduce this power loss, the hydraulic pump capacity is insufficient when the engine is running at low speed or when the vehicle is stopped, and the hydraulic actuator is There is a problem that a sufficient amount of hydraulic oil cannot be supplied.
[0003]
Thus, in hydraulic pumps for automobiles, it is required to reduce unnecessary power loss of the engine serving as a drive source while securing a sufficient amount of hydraulic oil when necessary. In the middle of the transmission system from the engine to the rotor, a person performs a switching operation to the speed change means by the action of a speed change means that can be switched between two levels of high and low, and a hydraulic pressure that is generated in the discharge oil passage according to the rotation of the rotor. And when the steering assisting hydraulic cylinder is operated by steering the steering wheel, the transmission means is switched to the high speed side by the action of the hydraulic pressure in the oil feed passage that rises along with the steering cylinder. Japanese Patent Application No. 10-10770 proposed a hydraulic pump unit configured to switch the speed change means to the low speed side when the hydraulic cylinder is not operated.
[0004]
However, in the power steering apparatus in which the hydraulic pump unit configured in this way is incorporated, when the transmission means is switched from the low speed side to the high speed side by the operation of the switching clutch, the hydraulic pump unit is operated accordingly. The rotation speed suddenly changes, and the oil pressure in the oil supply passage suddenly rises. As a result, the hydraulic oil from the hydraulic pump is switched based on the steering and supplied to the hydraulic cylinder from the servo valve. There is a problem that the flow rate changes suddenly, the steering assist force generated by the hydraulic cylinder changes suddenly, the steering wheel operation feeling changes suddenly, and the driver holding the steering wheel feels uncomfortable.
[0005]
In order to solve such a problem, the applicant of the present application improved the one proposed in Japanese Patent Application No. 10-10770, and a branching portion for branching the middle of the oil supply passage communicating with the hydraulic pump into a drain oil passage; A throttle means for generating a differential pressure corresponding to the supply flow rate to the hydraulic cylinder, and the branching portion, and opening and closing the drain oil passage by the action of the differential pressure to control the supply flow rate A hydraulic pump unit with a flow control valve was proposed in Japanese Patent Application No. 10-126470.
[0006]
In this Japanese Patent Application No. 10-126470, when the hydraulic cylinder is not operated without being steered, the flow control valve opens the drain oil passage, and a part of the pressure oil sent from the hydraulic pump is branched. When the hydraulic cylinder is operated by steering, the hydraulic pressure inside the oil supply path is increased, and the speed change means is switched to the high speed side by the operation of the switching clutch. The flow control valve operates in response to an increase in the differential pressure before and after the throttle means, and the drain oil passage is closed, and the oil flow to the drain oil passage is gradually reduced to supply the hydraulic cylinder. Gradually increases and the sudden change in the supply hydraulic pressure can be suppressed, so that the uncomfortable feeling can be reduced.
[0007]
[Problems to be solved by the invention]
By the way, in the case where the rotational speed of the hydraulic pump is switched between a high speed and a low speed as described above, the hydraulic pump rotates at a low speed when not being steered, and from the supply flow rate sent from the hydraulic pump to the hydraulic cylinder. However, since the drain oil flow rate to the drain oil passage is larger, the input torque-pressure characteristic line draws a curved line.
[0008]
Therefore, as previously proposed by the applicant of the present application, the one configured to operate the switching clutch by the action of the hydraulic pressure in the oil feed passage that rises with steering is steered from the steering neutral point, Steering until the hydraulic oil is operated by steering from the steering neutral point until the steering angle until the pressure oil in the oil supply passage rises to a predetermined value necessary for operating the switching clutch increases to a predetermined value or more. Since the angle becomes large, there is a concern that the steering becomes too heavy in the initial stage of steering.
[0009]
  The present invention has been made in view of such circumstances, and a throttle portion is provided in an oil discharge hole and an oil supply hole of a servo valve to which pressure oil is supplied from a hydraulic pump through a flow rate control valve.RukoAs a result, when steering starts from a steering neutral point with a small flow rate, the oil discharge flow rate through the oil discharge hole and the oil supply hole communicating with the oil discharge hole can be reduced early, and the hydraulic pressure from the servo valve can be reduced. Power that can increase the supply flow rate to the cylinder at an early stage, increase the hydraulic pressure to a predetermined value at an early stage, reduce the steering angle until the hydraulic cylinder operates in the initial stage of steering, and lighten the steering An object is to provide a steering device and a servo valve.
[0010]
[Means for Solving the Problems]
  A power steering apparatus according to a first aspect of the present invention includes a hydraulic pump, a hydraulic cylinder for assisting steering that communicates with the hydraulic pump via an oil feeding path, a flow control valve provided in the oil feeding path, and steering. And a servo valve for switching and supplying pressure oil from the flow control valve to the chamber of the hydraulic cylinder, the servo valve being fitted to the valve body so as to be capable of relative angular displacement, and the servo spool In the power steering apparatus in which a plurality of oil supply chambers, first intermediate pressure chambers, second intermediate pressure chambers, and first intermediate pressure chambers communicating with the oil supply holes are provided in this order between the valve bodies. At least one intermediate pressure chamber communicates with the oil drain hole and at least one otherSqueezedCommunicating with the oil drain holePleaseAndThe apertureThe first intermediate pressure chambers adjacent to both sides of the second intermediate pressure chamber communicating with the oil discharge hole having the groove portion communicate with the oil supply hole having the throttle portion.PleaseThe remaining first intermediate pressure chambers excluding the first intermediate pressure chamber communicate with the oil feed holes.
[0011]
  A power steering apparatus according to a second aspect of the present invention includes a hydraulic pump that is linked to a drive source via a transmission means, a steering assist hydraulic cylinder that is in communication with the hydraulic pump via an oil feed path, and the oil feed path. A flow control valve provided on the servomotor, a servo valve that supplies pressure oil from the flow control valve to the chamber of the hydraulic cylinder by switching based on steering, and a switch that causes the transmission to perform a shifting operation based on steering The servo valve includes a valve spool that is fitted to the valve body so as to be capable of relative angular displacement, and an oil supply chamber that communicates with the oil supply hole, a first intermediate pressure chamber, a second valve chamber, and the valve body. In the power steering apparatus in which a plurality of intermediate pressure chambers and a plurality of first intermediate pressure chambers are provided in this order, at least one of the second intermediate pressure chambers communicates with the oil discharge hole, and at least one otherSqueezedCommunicating with the oil drain holePleaseAndThe apertureThe first intermediate pressure chambers adjacent to both sides of the second intermediate pressure chamber communicating with the oil discharge hole having the groove portion communicate with the oil supply hole having the throttle portion.PleaseThe remaining first intermediate pressure chambers excluding the first intermediate pressure chamber communicate with the oil feed holes.
[0012]
  A servo valve according to a third aspect of the present invention includes a valve spool that is fitted to the valve body so as to allow relative angular displacement, and an oil supply chamber that communicates with the oil supply hole, a first intermediate pressure chamber, and a second intermediate pressure between the valve body. In the servo valve in which a plurality of chambers and a plurality of first intermediate pressure chambers are provided in this order, at least one of the second intermediate pressure chambers communicates with the oil drain hole and at least one otherSqueezedCommunicating with the oil drain holePleaseAndThe apertureThe first intermediate pressure chambers adjacent to both sides of the second intermediate pressure chamber communicating with the oil discharge hole having the groove portion communicate with the oil supply hole having the throttle portion.PleaseThe remaining first intermediate pressure chambers excluding the first intermediate pressure chamber communicate with the oil feed holes.
[0013]
In the first invention, the second invention and the third invention, since the magnitude of the load applied to the hydraulic cylinder is small when not being steered, the pressure oil supplied to the servo valve via the flow control valve is The entire flow rate is returned to the oil tank or the like from the oil drain hole having the throttle portion and the oil drain hole not having the throttle portion. When steering starts from the steering neutral point at this small flow rate, the load applied to the hydraulic cylinder gradually increases, the supply flow rate from the flow control valve to the servo valve gradually increases, and the valve spool of the servo valve rotates. The valve opening gradually increases and the flow rate supplied to the servo valve gradually increases.
[0014]
  At this timeThe secondAt least one of the two intermediate pressure chambers communicates with the oil drainage hole, and at least one of the other intermediate pressure chambers communicates with the oil drainage hole having the throttle portion, and the second intermediate pressure chamber communicates with the oil drainage hole having the throttle portion. The first intermediate pressure chambers adjacent to each other communicate with the oil feed hole having the throttle portion.For,Oil discharge flow rate by oil discharge holes and oil supply holes communicating with the oil discharge holes compared to those equipped with an existing servo valve in which all oil discharge holes have the same opening area and oil supply hole opening areas Can be reduced at an early stage, the supply flow rate from the servo valve to the hydraulic cylinder can be increased at an early stage, and the oil pressure can be raised to a predetermined value at an early stage. Therefore, the steering angle from when the hydraulic cylinder is steered at the small flow rate until the hydraulic cylinder starts operating can be reduced, and the steering can be lightened at the initial stage of steering.
[0015]
In the first and second inventions, the flow rate increases according to the valve characteristics of the flow control valve after the hydraulic pressure of the pressure oil supplied from the servo valve to the hydraulic cylinder rises to a predetermined value. Therefore, when the hydraulic cylinder starts to operate, the discontinuous portion of the oil pressure can be eliminated, and it is possible to eliminate a feeling of strangeness for the driver who is steering.
[0016]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, the present invention will be described in detail with reference to the drawings illustrating embodiments thereof.
Embodiment 1
FIG. 1 is a hydraulic system diagram of the power steering apparatus, and FIG. 2 is a sectional view of a switching portion for switching the rotational speed of the hydraulic pump.
[0017]
The power steering apparatus according to the present invention includes a hydraulic pump 3 that is linked to a drive source 1 that is an engine of an automobile via a speed change means 2, and a steering assist that communicates with a discharge portion of the hydraulic pump 3 via an oil feed path 4. And a flow rate control valve 6 that is provided in the middle of the oil feed path 4 and controls the flow rate of the pressure oil sent to the hydraulic cylinder 5 according to the magnitude of the load applied to the hydraulic cylinder 5; A servo valve 7 that supplies pressure oil from the flow rate control valve 6 to the chamber of the hydraulic cylinder 5 based on steering, and a switching clutch 8 that causes the transmission means 2 to perform a speed change operation based on steering. I have.
[0018]
The speed change means 2 is, for example, a planetary gear transmission, a planetary roller transmission, or the like that changes the rotational speed of the hydraulic pump 3 between, for example, a high speed and a low speed.
[0019]
In the flow control valve 6, a spool chamber 60 having a circular cross section is formed in a manner intersecting with the oil supply passage 4, and the flow control valve 6 is configured inside the spool chamber 60.
[0020]
  FIG. 3 is a sectional view showing a schematic configuration of the flow control valve 6 portion. The spool chamber 60 isFIG.As shown in FIG. 3, it is provided for the purpose of guiding the switching clutch 8 to the pressure oil chamber 9.LedOn the upstream side of the oil passage 10, one side in the longitudinal direction is formed to communicate with the middle of the oil feed passage 4. The other side of the spool chamber 60 isAnd minutesA throttle means provided in the middle of the oil feeding passage 4 is communicated with the bifurcated portion, more specifically, a downstream side of the fixed throttle 61 via a pressure guide passage 62. Further, an intermediate portion of the spool chamber 60 is The fluid is communicated with a suction oil passage 11 (see FIG. 2) of the hydraulic pump 3 maintained at a low pressure via a drain oil passage 63 opened at a position.
[0021]
The flow control valve 6 includes a first spool 6a having a bottomed cylindrical shape that is slidably inserted in the axial direction in the spool chamber 60 as described above, and an inner side of the first spool 6a. And a second spool 6b having a cylindrical shape that is slidably inserted in the axial direction. A through hole having a predetermined inner diameter is formed in the bottom of the first spool 6a, and a connecting rod 6c that projects coaxially from one side of the second spool 6b is formed in the through hole. A pressure receiving plate 6d having an outer diameter equal to that of the second spool 6b is coaxially fixed to the end of the connecting rod 6c that is fitted and protrudes to the outside of the first spool 6a.
[0022]
The first spool 6a and the second spool 6b are integrally energized in a direction away from each other by an energizing spring 64 interposed between the inner bottom surface of the former and the latter end surface opposite to the inner bottom surface. Are mounted inside the spool chamber 60 with the opening side of the first spool 6a directed to the communication side with the oil feed path 4 and the bottom side directed to the communication side with the pressure guiding path 62; The control spring 65 interposed between the pressure receiving plate 6d and the end face of the spool chamber 60 facing the pressure receiving plate 6d is urged toward the direction away from the end face, that is, toward the communication side with the oil feed path 4. Then, the opening side peripheral portion of the first spool 6 a is positioned so as to block the drain oil passage 63.
[0023]
That is, the inside of the spool chamber 60 faces the opening side of the first spool 6a, faces the first pressure chamber 60a communicated with the oil feeding path 4, and the bottom side of the first spool 6a, and The pressure guide passage 62 is divided into a second pressure chamber 60b that communicates with the drain oil passage 63. The drain oil passage 63 is blocked from communicating with the first pressure chamber 60a by the opening-side peripheral portion of the first spool 6a. It is in the state.
[0024]
Further, the arrangement chamber of the biasing spring 64 inside the first spool 6a is communicated with the opening of the drain oil passage 63 by a through hole 66 penetrating the peripheral surface of the first spool 6a. The sliding of the second spool 6b, which will be described later, with respect to the spool 6a is performed without any trouble due to the oil entering and exiting through the through hole 66.
[0025]
4 and 5 are explanatory views of the operation of the flow control valve 6 configured as described above. The pressure oil discharged by the operation of the hydraulic pump 3 flows into the first pressure chamber 60a on one side of the spool chamber 60 through the oil feed passage 4 as indicated by arrows in these drawings, The oil is fed to the hydraulic cylinder 5 through the fixed throttle 61 disposed in the middle of the oil feeding passage 4 on the downstream side.
[0026]
At this time, the hydraulic pressure P 1 is introduced into the first pressure chamber 60 a, and the second pressure chamber 60 b on the other side of the spool chamber 60 is downstream of the fixed throttle 61 via the pressure guiding path 62. Side hydraulic pressure P2 is introduced. The hydraulic pressure P2 is generated by reducing the pressure oil that has the hydraulic pressure P1 and is sent out to the oil feed passage 4 by the passage of the fixed throttle 61, and this reduced pressure amount is the amount of pressure oil that passes through the fixed throttle 61, that is, the hydraulic pressure. Since the pressure increases as the supply flow rate to the cylinder 5 increases, a differential pressure ΔP (= P1-−) increases between the first pressure chamber 60a and the second pressure chamber 60b as the supply flow rate increases. P2) occurs.
[0027]
Here, the second spool 6b receives the hydraulic pressure P1 of the first pressure chamber 60a at one end face exposed on the opening side of the first spool 6a, and the hydraulic pressure P2 of the second pressure chamber 60b is integrated on the same side. The pressure receiving plate 6d receives pressure on the end face of the pressure receiving plate 6d. As described above, the pressure receiving plate 6d has an outer diameter substantially equal to that of the second spool 6b, and the first pressure chamber 60a and the second pressure chamber. The pressure receiving area on the 60b side is the same. Accordingly, the second spool 6b is pressed from the first pressure chamber 60a toward the second pressure chamber 60b as the differential pressure ΔP increases, and the force balance between the pressing force and the spring force of the control spring 65 is achieved. Accordingly, it moves toward the second pressure chamber 60b.
[0028]
On the other hand, the first spool 6a receives the hydraulic pressure P1 of the first pressure chamber 60a on the peripheral edge on the opening side, and receives the hydraulic pressure P2 of the second pressure chamber 60b on the bottom surface. Is bigger. Therefore, the first spool 6a is not driven by the differential pressure ΔP but by the action of the hydraulic pressure P2 of the second pressure chamber 60b applied to the bottom surface side, that is, the hydraulic pressure P2 downstream of the fixed throttle 61, from the second pressure chamber 60b. It is pressed toward the first pressure chamber 60a, and is directed toward the first pressure chamber 60a with respect to the second spool 6b in accordance with the force balance between the pressing force and the spring force of the biasing spring 64. Move relative.
[0029]
When the hydraulic pressure P2 generated downstream of the fixed throttle 61 is sufficiently small, such as when the hydraulic cylinder 5 is not operated without being steered, the hydraulic pressure P1 of the first pressure chamber 60a, the hydraulic pressure P2 of the second pressure chamber 60b, Since the differential pressure ΔP between them is small, the first spool 6a and the second spool 6b move integrally leftward from the assembly position shown in FIG. Is slightly opened, and most of the pressure oil delivered from the hydraulic pump 3 into the oil feed passage 4 is fed to the hydraulic cylinder 5 through the fixed throttle 61.
[0030]
When the hydraulic cylinder 5 is operated by steering from this state and the hydraulic pressure P2 increases, the differential pressure ΔP before and after the fixed throttle 61 also increases. As a result, the second spool 6b is pressed by the action of the differential pressure ΔP. Exceeds the spring force of the control spring 65, and starts to move integrally with the outer first spool 6a against this spring force. As shown in FIG. The opening area of the drain oil passage 63 to the inside becomes large, and most of the supply flow rate supplied from the hydraulic pump 3 into the oil feed passage 4 is drained to the low pressure side through the drain oil passage 63.
[0031]
The oil discharge flow rate at this time increases as the differential pressure ΔP before and after the fixed throttle 61 increases, that is, as the supply flow rate increases, so the supply flow rate to the hydraulic cylinder 5 increases the rotation speed of the hydraulic pump 3. Regardless, it remains almost constant.
[0032]
On the other hand, as described above, the first spool 6a pressed by the action of the hydraulic pressure P2 in the second pressure chamber 60b has a pressing force by the hydraulic pressure P2 exceeding the spring force of the biasing spring 64 and the second spool 6b. The relative movement described above is started. As a result, as shown in FIG. 5, the opening area of the drain oil passage 63 to the inside of the spool chamber 60 is reduced, and the flow rate of oil discharged to the low pressure side through the drain oil passage 63 is reduced. The supply flow rate to 5 increases as the hydraulic pressure P2 increases.
[0033]
FIG. 6 is a flow characteristic diagram showing the relationship between the pressure of the downstream pressure oil of the flow control valve 6 and its flow rate. As shown in the figure, the amount of oil supplied from the hydraulic pump 3 is kept substantially constant as shown in (a) while the downstream side hydraulic pressure P2 is small and the state change shown in FIGS. While the oil pressure P2 exceeds the predetermined limit pressure Ps and the state change shown in FIGS. 4 to 5 occurs in the flow control valve 6, the pressure increases as the oil pressure P2 increases. When P2 rises to a predetermined oil pressure P, the downstream flow rate returns to a large flow rate as shown in FIG. Accordingly, by setting the limit pressure PS so as to substantially coincide with the rotational speed at which the speed change means 2 is switched to the high speed side by the operation of the switching clutch 8, it is possible to prevent a sudden change in hydraulic pressure occurring before and after this switching. it can. The above-described setting of the limit pressure PS can be realized by appropriately selecting the pressure receiving area difference on both sides of the first spool 6a and the spring force of the biasing spring 64.
[0034]
The servo valve 7 includes a valve spool 7a connected to the steered wheel via an input shaft, and a valve body 7b inserted around the valve spool 7a and connected to the output shaft. The basic configuration of the valve spool 7a and the valve body 7b is the same as the existing one, and in the first embodiment, eight equally spaced servo valves are used. On the inner peripheral surface of the cylindrical valve body 7b, eight first oil grooves 70, each having an equal width, are arranged in parallel in the circumferential direction, and the inner diameter of the valve body 7b Similarly, on the outer peripheral surface of the thick cylindrical valve spool 7a having substantially the same outer diameter, eight second oil grooves 71, each having an equal width, are arranged in parallel in a circumferential direction. ing.
[0035]
The valve spool 7a is fitted inside the valve body 7b so as to be capable of relative rotation on the same axis, and both are connected to each other by a torsion bar 12 inserted inside the valve spool 7a. The first oil grooves 70 and the second oil grooves 71 are arranged in a zigzag manner in the circumferential direction as shown in the figure in a neutral state where the torsion bar 12 is not twisted, and are adjacent to each other on both sides. Positioned to communicate with
[0036]
With the above configuration, each of the first oil grooves 70 of the valve body 7b faces the land between the second oil grooves 71 of the valve spool 7a, and the second oil grooves 71 of the valve spool 7a. Are opposed to the lands between the first oil grooves 70 of the valve body 7b, and on the inner periphery of the first oil grooves 70 on the fitting circumference of the valve body 7b and the valve spool 7a ( The eight oil chambers between the groove edges on both sides in the width direction and the eight oil chambers on the outside (between the groove edges on both sides in the width direction) of the second oil grooves 71 have communication portions therebetween. Are alternately arranged.
[0037]
The valve body 7b and the valve spool 7a are capable of relative angular displacement within the torsional range of the torsion bar 12 connecting them, and the communication portions between the oil chambers, that is, the oil grooves 70, 71,. The groove edges on both sides in the width direction act as throttle portions a and b that increase / decrease the respective communication areas (throttle areas) according to the relative angular displacement.
[0038]
Of the eight oil chambers formed by the second oil grooves 71 of the valve spool 7a, four of the other oil chambers, which are located every other, penetrate the peripheral wall of the valve body 7b and are located outside the respective oil grooves 71. An oil supply chamber 73, which is connected to the downstream side of the flow rate control valve 6 via the fixed throttle 61 via separate oil supply holes 72 having openings, is supplied with pressure oil from the hydraulic pump 3. ing. On the other hand, the remaining four oil chambers penetrate the valve spool 7a in the radial direction, and each oil drain hole 74 has an opening at the bottom of each oil groove 71, and a hollow portion inside the valve spool 7a. Are connected to an oil tank 13 serving as an oil discharge destination, thereby constituting second intermediate pressure chambers (hereinafter referred to as oil discharge chambers) 75... Serving as a passage for discharged oil to the oil tank 13.
[0039]
On the other hand, of the eight oil chambers formed on the inner side of the first oil grooves 70, the four oil chambers adjacent to the oil supply chamber 73 on the same side in the circumferential direction are formed on the peripheral wall of the valve body 7b. This cylinder chamber is connected to one cylinder chamber 5R of the hydraulic cylinder 5 to which pressure oil is fed, through each of the oil feed holes 76 which pass through and have an opening at the bottom of each oil groove 70. 5R first intermediate pressure chambers (hereinafter referred to as first oil feeding chambers) 77... And the remaining four are the other cylinders of the hydraulic cylinder 5 through similar oil feeding holes 78. A first intermediate pressure chamber (hereinafter referred to as a second oil feeding chamber) 79... Is connected to the chamber 5L and connected to the cylinder chamber 5L. Accordingly, oil passages are formed on both sides of the oil supply chambers 73 through the first oil supply chamber 77 or the second oil supply chamber 79 to the oil discharge chamber 75, respectively. The oil discharge chamber 75 and the oil feed chamber 79 communicate with each other through the throttle portion a, and the oil supply chamber 73 and the oil feed chamber 79 and the oil discharge chamber 75 and the oil feed chamber 77 communicate with each other through the throttle portion b. The
[0040]
The servo valve 7 configured as described above is a basic configuration of an existing eight-equal distribution type. In the present invention, as described above, the throttle portion 80 is provided every other even number of oil drain holes 74, and each oil supply is provided. The first oil drain hole 74a having the throttle part 80 on one side with respect to the holes 72, and the second oil drain hole 74b not having the throttle part 80 on the other side are arranged, respectively. The throttle parts 81, 81 are connected to the oil feed holes 76, 78 communicating with the first and second oil feed chambers 77, 79 adjacent to each other on both sides of the oil drain chamber 75 communicating with the first oil drain hole 74 a having the portion 80. Are provided as first oil feed holes 76a and 78a, and the first and second oil feed chambers adjacent to each other on both sides of the oil drain chamber 75 communicating with the second oil drain hole 74b having no throttle 80 are provided. No throttle part is provided in the oil feed holes 76, 78 communicating with 77, 79, and this is used as the second oil feed holes 76b, 78b. The oil discharge flow rate of the pressure oil supplied to the oil supply holes 72 is the oil discharge flow rate through the first and second oil discharge holes 74a and 74b and the oil feed holes 76a and 76b communicating with the oil discharge holes 74a and 74b. Alternatively, the average oil discharge flow rate obtained by averaging the oil feed holes 78a and 78b is set.
[0041]
The throttle portions 80 and 81 are made of a cylindrical member such as a pipe. The throttle portion 80 is inserted into the oil drain hole 74 of the existing servo valve 7, and the first oil drain hole 74a provided with the throttle portion 80 is provided. The oil drainage flow rate is reduced compared to the oil drainage flow rate of the second oil drainage hole 74b in which the throttle portion 80 is not provided. Further, the throttle part 81 is inserted into the oil feed hole 76 or the oil feed hole 78 of the existing servo valve 7, and the oil discharge flow rate of the first oil feed holes 76 a and 78 a provided with the throttle part 81 is changed to the throttle part. Compared to the oil discharge flow rate of the second oil feed holes 76b, 78b where 81 is not provided.
[0042]
FIG. 7 is a hydraulic characteristic diagram showing the relationship between the steering angle and the opening area of the oil discharge hole and oil supply hole of the servo valve 7. The broken line (a) is the configuration of the present invention, and (b) and ( The solid line a in a) is an existing configuration, and the solid lines b in (c) and (a) are provided with throttle portions 80 in all of the oil drain holes of the existing configuration, and the throttle portions 81 in all of the oil feed holes. This is a configuration provided. When all the oil drain holes 74 and all the oil feed holes 76 and 78 do not have the throttle portions 80 and 81, the amount of reduction in the opening area at the initial stage of steering is as shown by the solid line a in FIGS. 7 (b) and (a). When the throttle portions 80 and 81 are provided in all the oil drain holes 74 and the oil feed holes 76 and 78, the initial stage of steering as shown by the solid line B in FIGS. 7 (c) and (a). The amount of reduction in the opening area is constant and does not change at all. On the other hand, at least one oil drain hole 74 is provided with a throttle part 80 to form first and second oil drain holes 74a and 74b. The throttle portions 81, 81 are provided in the oil feed holes 76, 78 communicating with the first and second oil feed chambers 77, 79 adjacent to both sides of the oil drain chamber 75 communicating with the first oil drain hole 74 a having 80. In the present invention, the first oil feeding holes 76a and 78a and the second oil feeding holes 76b and 78b are provided. As shown by the broken line in (a), (b) and (c) can be changed to an averaged opening area.
[0043]
FIG. 8 is a hydraulic characteristic diagram showing the relationship between the steering angle and the hydraulic pressure controlled by the servo valve.
In this way, by changing the drainage holes 74a and 74b and the oil feed holes 76a and 76b or the oil feed holes 78a and 78b to all the oil feed holes 72 to an averaged opening area, the flow rate control valve 6 can reduce the flow rate on the downstream side. When the steering angle required for the oil pressure to rise to the predetermined oil pressure P is θ, the oil drain holes 74a and 74b and the oil feed holes 76a and 76b or the oil feed holes 78a and 78b when the steering angle becomes θ. When the steering angle is less than θ degrees, oil is discharged under the influence of the throttle portions 80 and 81, and a hydraulic characteristic corresponding to a small flow rate is obtained. When the steering angle is greater than θ degrees, the oil is discharged without being influenced by the throttle portions 80 and 81, and the hydraulic characteristics of the existing 8-equal servo valve, that is, a large flow rate as shown in FIG.
[0044]
9 and 10 are explanatory views of the operation of the servo valve 7 when the steering is started.
For example, as shown in FIG. 9, when the steering wheel is steered rightward from the steering neutral point, the valve spool 7a rotates clockwise with respect to the valve body 7b, the throttle part a is gradually opened, and the throttle part b is gradually closed. The oil supply holes 78a and 78b are directly drained from the oil drain holes 74a and 74b arranged on both sides of all the oil feed holes 72 and are arranged counterclockwise with respect to all the oil feed holes 72. Since the oil is supplied from the oil drain holes 74a and 74b to the oil feed holes 76a and 76b, the steering angle becomes θ by changing the throttle amount by the throttle parts a and b. The opening areas of the oil drain holes 74a and 74b and the oil feed holes 78a and 78b can be made closer to the opening areas of the throttle portions 80 and 81.
[0045]
Further, as shown in FIG. 10, when the steering wheel is steered leftward from the steering neutral point, the valve spool 7a rotates counterclockwise with respect to the valve body 7b, the throttle part b is gradually opened, and the throttle part a is gradually increased. The oil supply holes 76a and 76b are closed and drained directly from the oil drain holes 74a and 74b arranged on both sides of all the oil feed holes 72, and are arranged in the clockwise direction with respect to all the oil feed holes 72. Since the oil is discharged from the oil drain holes 74a and 74b through the oil and supplied to the oil feed holes 78a and 78b, the steering angle becomes θ by changing the throttle amount by the throttle parts a and b. The opening areas of the oil drain holes 74a and 74b and the oil feeding holes 76a and 76b can be made close to the opening areas of the throttle portions 80 and 81.
[0046]
Next, the operation of the power steering apparatus configured as described above will be described.
When the steering wheel is not being steered, such as when idling, the load applied to the hydraulic cylinder 5 is small, so that the transmission means 2 is switched to the low speed rotation side by the switching clutch 8, and the hydraulic pump 3 is at low speed rotation. The flow rate of the pressure oil supplied from the pump 3 to the oil feed path 4 is small as shown in FIG. Therefore, the pressure oil supplied to the servo valve 7 via the flow control valve 6 has a total flow rate from the first oil drain hole 74 a having the throttle part 80 and the second oil drain hole 74 b not having the throttle part 80. The oil is returned to the oil tank 13 and the like.
[0047]
The flow rate of oil introduced from the oil supply hole 72 to the oil supply chamber 73 is a small flow rate, and the pressure oil of this small flow rate is equally distributed, so that the pressure oil distributed to the four oil supply chambers 73 is 4 The first and second oil distribution chambers are equally distributed to the oil passages on both sides of the oil supply chamber 73, reach the oil discharge chamber 75 via the first oil supply chamber 77 and the second oil supply chamber 79, and open to these. The oil flows into the hollow portion inside the valve spool 7a through the oil drain holes 74a and 74b, joins in the hollow portion, and is discharged into the oil tank 13. Accordingly, there is no pressure difference between the oil feeding chambers 77 and 79 and between the cylinder chambers SR and SL of the hydraulic cylinder 5 connected to each of them, and the hydraulic cylinder 5 does not generate any force.
[0048]
When steering is performed from the steering neutral point in this small flow rate state, the magnitude of the load applied to the hydraulic cylinder 5 gradually increases, and the flow rate of the pressure oil supplied to the servo valve 7 by the flow rate control valve 6 is accordingly increased. As the valve spool 7a of the servo valve 7 rotates as shown in FIG. 9 or FIG. 10, the valve opening gradually increases and the flow rate supplied to the oil supply hole 72 gradually increases. At this time, the throttle part 80 is provided in at least one oil drain hole 74, and the throttle part 81 is provided in a pair of oil feed holes 76 and 78 disposed on both sides of the oil drain hole 74a having the throttle part 80. Since each of the oil drain holes 74 is provided, the opening area of all the oil drain holes 74 and the opening area of all the oil feed holes 76 and 78 are the same as those provided with the existing servo valve. The oil discharge flow rate can be reduced early, the supply flow rate to the oil feed hole 76 or the oil feed hole 78 can be increased early, and the oil pressure is set to a predetermined value Ps necessary for operating the hydraulic cylinder 5. It can be raised early.
[0049]
Therefore, the steering angle from when the hydraulic cylinder 5 is steered from the steering neutral point in a small flow rate state until the hydraulic cylinder 5 starts operating can be reduced, and the steering can be lightened at the initial stage of steering.
[0050]
Thus, when the oil pressure in the oil feed hole 76 or the oil feed hole 78 rises to the predetermined value PS, the switching clutch 8 is operated, the transmission means 2 is switched to the high speed side, and the above-described operation of the flow control valve 6 is performed. Therefore, the hydraulic pressure of the pressure oil supplied to the hydraulic cylinder 5 is asymptotically continuously from the change state at the rotational speed before switching as shown in FIG. 6B to the change state at the rotational speed after switching. The discontinuity can be eliminated. In addition, since the steering angle from when the hydraulic cylinder 5 starts to operate after being steered from the steering neutral point in a small flow rate state, the transition length H from the small flow curve A to the multiple flow curve B is applied to the present invention. The steering assist force generated by the hydraulic cylinder 5 can be shortened compared to the transition length H1 previously proposed by the person, and does not change suddenly before and after the predetermined value Ps, so that the driver who is steering is uncomfortable. Never give.
[0051]
Embodiment 2(Reference example)
  FIG. 11 is a hydraulic system diagram of the power steering apparatus.
  The power steering apparatus of the second embodiment and the servo valve 7 used therefor areReference exampleInstead of providing the throttle portion 81 in the oil feed holes 76 and 78, the oil feed of a pair of oil feed chambers 77 and 79 adjacent on both sides of the oil drain chamber 75 communicating with the oil drain hole 74a having the throttle portion 80 is provided. The communication with the holes 76 and 78 is cut off, in other words, the oil feed holes 76a and 78a of the first embodiment are eliminated, and the exhaust oil flow rates of the pressure oil supplied to all the oil supply holes 72 are the first and second. The oil discharge flow rate of the oil discharge holes 74a and 74b is an average oil discharge flow rate, and other configurations and operations are the same as those in the first embodiment. The detailed description and structure, operation, and effect are omitted.
[0052]
The communication with the oil supply holes 76 and 78 of the oil supply chambers 77 and 79 is interrupted by, for example, not forming at least one set of oil supply holes 76 and 78 in the existing valve body 7b, The buried body is inserted into the oil supply holes 76 and 78 adjacent to both sides of the oil discharge chamber 75 communicating with the oil discharge hole 74a formed in the valve body 7b and having the throttle portion 80 and closed.
[0053]
In the second embodiment, the pressure oil supplied to all of the oil supply holes 72 is a set of oil supply adjacent to both sides of the oil discharge chamber 75 communicating with the oil discharge hole 74a having the throttle portion 80. The pressure oil supplied from all of the chambers 77 and 79 to the oil discharge holes 74a and 74b through the oil supply holes 76 and 78 and not supplied to all the oil supply holes 72 is directly discharged to the oil discharge holes 74a and 74b. Thus, the same effect as that of the first embodiment can be obtained.
[0054]
Embodiment 3(Reference example)
  FIG. 12 is a hydraulic system diagram of the power steering apparatus.
  The power steering apparatus according to the third embodiment and the servo valve 7 used therefor areReference exampleInstead of providing the throttle 80 in the drain hole 74, the communication with the drain hole 74 of at least one of the drain chambers 75 is blocked, in other words, the drain provided with the throttle 80 in the first embodiment. A pair of oil feeding chambers 77 and 79 adjacent on both sides of the oil draining chamber 75 that has eliminated the oil hole 74a and cut off the communication with the oil draining hole 74 are connected to the oil feeding holes 76a and 78a having the throttle portion 81. The oil discharge flow rate of the pressure oil that is communicated and supplied to all the oil supply holes 72 is an average obtained by averaging the oil discharge flow rates of the oil discharge holes 74 and the oil supply holes 76a and 76b or the oil supply holes 78a and 78b. The oil flow rate is the same, and the other configurations and operations are the same as those in the first embodiment. Therefore, common parts are denoted by the same reference numerals, and detailed descriptions and structures, operations, and effects are omitted. To do.
[0055]
The communication with the oil drain hole 74 of the oil drain chamber 75 is interrupted by, for example, not drilling at least one oil drain hole 74 in the existing valve spool 7a, or drilling in the existing valve spool 7a. The buried body is inserted into the oil drain hole 74 between the oil feed chambers 77 and 79 communicating with the oil feed holes 76a and 78a having the throttled portion 81 and closed.
[0056]
In the third embodiment, the pressure oil supplied to all of the oil supply holes 72 is directly discharged from the oil discharge holes 74 and the oil discharge chamber in which communication with the oil discharge holes 74 is blocked. As in the first embodiment, oil is discharged from a pair of oil supply chambers 77 and 79 adjacent to both sides of 75 to oil discharge holes 74 through oil supply holes 76a and 76b or oil supply holes 78a and 78b. The following effects can be obtained.
[0057]
Embodiment 4
Although the power steering apparatus of the fourth embodiment and the servo valve 7 used for the power steering apparatus are not shown, a throttle portion 80 formed separately from the valve spool 7a is formed in the oil drain hole 74 of the first and second embodiments. Instead of providing the throttle part 81 formed separately from the valve body 7b in the oil feed holes 76 and 78 of the first and third embodiments, the throttle part 80 is connected to the valve spool 7a. The throttle part 81 is formed integrally with the valve body 7b, and the other structures and operations are the same as those of the first, second, and third embodiments. Description, structure, operation and effect are omitted. In the fourth embodiment, an existing servo valve can be used at a low cost.
[0058]
The power steering apparatus according to the present invention includes, for example, a small flow rate type hydraulic pump and a large flow rate type hydraulic pump 3 instead of the speed change means 2, and a small flow rate when the steering wheel is not steered. When a hydraulic pump of a type is selected and steered from a steering neutral point and raised to a predetermined value necessary to operate the hydraulic cylinder 5, a high flow type hydraulic pump may be selected. In short, it is sufficient if the flow rate can be changed between a small flow rate and a large flow rate.
[0059]
In the first to fourth embodiments described above, the servo valve 7 has an equal number of 4 or more having a plurality of oil drain holes 74, for example, an even number of 6 equal, 12 equal, 14 equal, 16 equal, etc. It is preferable that the number of oil drain holes is preferably equal to or more than 8. However, a servo valve having an odd number of oil drain holes may be used. In short, at least one of the oil drain holes is an oil drain hole. 74, and at least one of the other fluids is communicated with an oil drain hole 74 a having a throttle portion 80 that blocks communication with the oil drain hole 74, and the oil drain chamber (second intermediate pressure chamber) 75 is shut off. Alternatively, a pair of oil feeding chambers (first intermediate pressure chambers) 77 and 79 adjacent to both sides of an oil draining chamber (second intermediate pressure chamber) 75 communicating with the oil draining hole having the throttle portion 81 are oil feeding holes 76, 79. The communication with 78 is communicated with oil supply holes 76a and 78a having a throttle part 81, and the rest excluding the oil supply chamber Okuaburashitsu may have a configuration in communication with the Okuaburaana 76,78.
[0060]
【The invention's effect】
As described above in detail, according to the first invention, the second invention, and the third invention, all the oil discharge holes have the same opening area, and all the oil feeding holes have the same opening area. Compared with a servo valve, the oil flow rate through the oil drain hole can be reduced early, and the supply flow rate from the servo valve to the hydraulic cylinder can be increased early, and the oil pressure can be increased to a predetermined value early. Therefore, the steering angle from when the hydraulic cylinder is steered at a small flow rate until the hydraulic cylinder starts operating can be reduced, and steering can be lightened at the initial stage of steering.
[0061]
Moreover, according to the first and second aspects of the invention, when the hydraulic cylinder starts to operate, the discontinuity of the oil pressure can be eliminated, and the driver performing the steering wheel operation can be prevented from feeling uncomfortable. It is.
[Brief description of the drawings]
FIG. 1 is a hydraulic system diagram of a first embodiment of a power steering apparatus according to the present invention.
FIG. 2 is a cross-sectional view of a switching portion for switching the rotational speed of the hydraulic pump of the power steering apparatus according to the present invention.
FIG. 3 is a cross-sectional view showing a schematic configuration of a flow control valve portion of the power steering apparatus according to the present invention.
FIG. 4 is an operation explanatory view of a flow control valve of the power steering apparatus according to the present invention.
FIG. 5 is an operation explanatory view of a flow control valve of the power steering apparatus according to the present invention.
FIG. 6 is a flow characteristic diagram showing the relationship between the pressure of the downstream pressure oil of the flow control valve of the power steering apparatus according to the present invention and the flow rate thereof.
FIG. 7 is a hydraulic characteristic diagram showing the relationship between the steering angle of the power steering apparatus according to the present invention and the opening area of the oil drain hole of the servo valve.
FIG. 8 is a hydraulic characteristic diagram showing the relationship between the steering angle of the power steering apparatus according to the present invention and the hydraulic pressure controlled by a servo valve.
FIG. 9 is an operation explanatory view of the servo valve when the steering of the power steering apparatus according to the present invention is started.
FIG. 10 is an operation explanatory view of the servo valve when the steering of the power steering apparatus according to the present invention is started.
FIG. 11Reference exampleIt is a hydraulic system figure of Embodiment 2 of the power steering device which concerns on.
FIG.Reference exampleIt is a hydraulic system figure of Embodiment 3 of the power steering device which concerns on.
[Explanation of symbols]
1 Drive source
2 Shifting means
3 Hydraulic pump
4 Oil supply passage
5 Hydraulic cylinder
6 Flow control valve
7 Servo valve
7a Valve spool
7b Valve body
72 Refueling hole
73 Refueling chamber
74 Oil drain hole
75 Oil drain chamber
76, 78 Oil feed hole
77, 79 Oil feeding chamber
8 switching clutch
80 Aperture
81 Aperture

Claims

A hydraulic pump, a hydraulic cylinder for assisting steering communicating with the hydraulic pump via an oil feed path, a flow control valve provided in the oil feed path, and pressure oil from the flow control valve based on steering A servo valve that is switched and supplied to the chamber of the hydraulic cylinder, and the servo valve communicates with the oil supply hole between the valve spool and the valve body that are fitted to the valve body to allow relative angular displacement. In the power steering apparatus in which a plurality of oil supply chambers, first intermediate pressure chambers, second intermediate pressure chambers, and first intermediate pressure chambers are provided in this order, at least one of the second intermediate pressure chambers is an oil discharge hole. communicates with, and communicates with the oil discharge hole having other at least one of grain Ri unit, first intermediate adjacent on both sides of the second intermediate-pressure chamber communicating with the oil discharge hole having the narrowed portion chamber is communicated with the oil feed hole having a throttle portion, first in Power steering apparatus and the remaining first intermediate pressure chamber, characterized in that in communication with the oil feed hole except chamber.

Based on a hydraulic pump linked to a drive source via a speed change means, a steering assist hydraulic cylinder communicating with the hydraulic pump via an oil feed path, a flow rate control valve provided in the oil feed path, and steering A servo valve that switches and supplies the pressure oil from the flow control valve to the chamber of the hydraulic cylinder, and a switching clutch that causes the shifting means to perform a shifting operation based on steering. The oil supply chamber, the first intermediate pressure chamber, the second intermediate pressure chamber, and the first intermediate pressure chamber that communicate with the oil supply hole are located between the valve spool and the valve body that are fitted to the valve body so as to allow relative angular displacement. in the power steering apparatus is provided with a plurality, in the order, the second intermediate pressure chamber communicates with the at least one oil discharge hole, communicates with the oil discharge hole having other at least one of aperture Ri unit and, an oil discharge hole having the narrowed portion The first intermediate-pressure chamber adjacent on both sides of the second intermediate-pressure chamber which communicates with and is communicated with the oil feed hole having a throttle portion, the remaining of the first intermediate-pressure chamber with the exception of the first intermediate pressure chamber oil feed A power steering apparatus characterized by communicating with a hole.

The oil supply chamber, the first intermediate pressure chamber, the second intermediate pressure chamber, and the first intermediate pressure chamber that communicate with the oil supply hole are located between the valve spool and the valve body that are fitted to the valve body so as to allow relative angular displacement. the servo valve that is present in a plural order, the second intermediate pressure chamber communicates with the at least one oil discharge bore is communicated with the oil discharge hole having other at least one of aperture Ri unit the first intermediate-pressure chamber adjacent on both sides of the second intermediate-pressure chamber communicating with the oil discharge hole having the narrowed portion is communicated with the oil feed hole having a throttle portion, the first intermediate-pressure chamber A servo valve characterized in that the remaining first intermediate pressure chamber communicates with the oil feed hole.