JP3821335B2 - Hydraulic booster and brake hydraulic booster system using this hydraulic booster - Google Patents

Description

【００６２】
で与えられるとともに、図４に示すように直線αで表される。
【００６３】
動力室２５内の液圧が更に上昇してパワーピストン８が更に前進し、各ホイールシリンダ２８,２９；５１,５２のロスストロークが解消すると、各ホイールシリンダ２８,２９；５１,５２がブレーキ力を発生し、実質的にブレーキが作動するようになる。この状態では、上昇した動力室２５内の液圧により反力ピストン２０の右端２０ｅが入力軸１８の段部１８ａに当接し、反力ピストン２０は動力室２５内の液圧による付勢力で入力軸１８に力を入力軸１８の入力に対抗するように作用する。したがって、入力軸１８に作用される反力が大きくなって、ジャンピング作用が終了し、以後、ブレーキ液圧倍力装置１の出力は入力軸１８の入力に対してロスストローク中よりは小さく上昇する。すなわち、ブレーキ液圧倍力装置１は反力が大きくなることから比較的小さなサーボ比で入力軸１８の入力を倍力して出力するサーボ制御を行うとともに、動力室２５内の液圧がこのサーボ比に対応した液圧となる。このときのサーボ比は通常ブレーキ時のサーボ比に設定されている。また、このサーボ制御中で、入力が所定の大きさになるまでは、動力室２５内の液圧が圧力切換弁４６の作動圧まで上昇しないので、圧力切換弁４６は第１位置Iに設定されたままとなり、反力室４１はリザーバ３３に接続されたままとなっている。
【００６４】
そして、各ホイールシリンダ２８,２９；５１,５２はそれぞれ入力軸１８の入力に対して倍力されたブレーキ力を発生し、このブレーキ力でブレーキが作動する。このとき、前述のように動力室２５内の液圧とマスタシリンダ圧とはバランスして互いに等しくなり、各ホイールシリンダ３３,３４；５８,５９が発生するブレーキ力も互いに等しくなる。このときの、ブレーキ液圧倍力装置１の入出力特性は、
【００６５】
【数２】

【００６６】
で与えられるとともに、図４に示すように直線βで表される。
【００６７】
入力が所定量になって、動力室２５内の液圧が圧力切換弁４６の作動圧になると、圧力切換弁４６は切り換えられて第２位置IIに設定される。すると、反力室４１は出力口２７とホイールシリンダ２８,２９との間のブレーキ液通路に接続され、反力室４１には、サーボ比制御圧である出力口２７の液圧、つまり動力室２５の液圧が導入される。そして、反力室４１に導入された液圧は、入力軸１８の段部１８ａに当接している反力ピストン２０の右端２０ｅの一部に、入力軸１８に加えられている入力と同方向に作用するようになる。このため、入力軸１８に作用される反力が小さくなって、直線βで表される通常ブレーキのサーボ比でのサーボ制御が終了し、以後、ブレーキ液圧倍力装置１の出力は入力軸１８の入力に対して通常ブレーキ時のサーボ制御中よりは大きく上昇する。すなわち、ブレーキ液圧倍力装置１は反力が小さくなることから比較的大きなサーボ比で入力軸１８の入力を倍力して出力するサーボ制御を行うようになる。このときのサーボ比は通常ブレーキ時のサーボ比より大きくなる。これにより、各ホイールシリンダ２８,２９；５１,５２はそれぞれ入力軸１８の入力に対して通常ブレーキ時のブレーキ力より大きなブレーキ力を発生する。このときの、ブレーキ液圧倍力装置１の入出力特性は、
【００６８】
【数３】

【００６９】
で与えられるとともに、図４に示すように直線γで表される。このように、ブレーキ倍力装置１は、入力が所定以上大きくなると通常ブレーキ時のサーボ比より大きなサーボ比でサーボ制御を行う逆二段サーボ特性を有している。
【００７０】
更に、入力が上昇して、動力室２５の液圧がアキュムレータ４０に蓄圧される最大設定圧になると、動力室２５の液圧はそれ以上上昇しなく、ブレーキ液圧倍力装置１は大きなサーボ比によるサーボ制御を終了し、全負荷状態となる。したがって、これ以後、ブレーキ液圧倍力装置１の出力上昇分は、入力上昇分を倍力しないものとなる。
【００７１】
ブレーキペダルを解放してブレーキ作動を解除すると、入力軸１８および弁作動部材１７がともに右方へ後退して制御弁６０の第２弁座１７ａが円錐弁１４から離座し、動力室２５内の圧液が、弁座部材１０の軸方向孔１０ｃ、円錐弁１４と第２弁座１７ａとの間の隙間、弁体１５の軸方向孔３０，径方向孔３１、段付孔４の小径部４ａ、および排出口３２を介してリザーバ３３に排出される。このとき、弁作動部材１７のストッパ部１７ｂが、反力ピストン２０のストッパ部２０ｃに当接するまで、入力軸１８が大きく後退するので、第２弁座１７ａが円錐弁１４から大きく開き、動力室２５内の圧液は迅速に排出される。
【００７２】
動力室２５内の圧液の排出により、一方の系統の両ホイールシリンダ２８,２９の圧液も迅速に動力室２５を通ってリザーバ３３に排出されて、両ホイールシリンダ２８,２９の液圧が低下する。一方、リターンスプリング５９のばね力により、マスタシリンダピストン４４およびパワーピストン８が迅速に後退するため、液室４９の液圧および他方の系統の両ホイールシリンダ５１,５２の液圧がともに低下する。そして、弁ロッド５５が弁開放ロッド５６に当接すると、それ以後のマスタシリンダピストン４７の後退に対して、弁５４が弁座５８から離座し、液室４９がリザーバ３３に接続される。このため、両ホイールシリンダ５１,５２の圧液も迅速に液室４９を通ってリザーバ３３に排出されて、両ホイールシリンダ５１,５２の液圧が更に低下する。これにより、両系統のブレーキが迅速に解除開始される。
【００７３】
動力室２５内の液圧が圧力切換弁４６の設定作動圧より低下すると、圧力切換弁４６が第１位置Iに切り換わり、反力室４１がリザーバ３３に接続される。動力室２５内の液圧が所定圧に更に低下すると、スプリング２１のばね力により反力ピストン２０がパワーピストン８および入力軸１８に対して相対的に前進して弁座部材１０のフランジ部１０ｂに当接するとともに、反力ピストン２０の右端２０ｅが入力軸１８の段部１８ａから離隔する。
【００７４】
ブレーキ解除がほぼ終了するまで入力軸１８が更に後退すると、弁作動部材１７のストッパ部１７ｂがプラグ６の小径突出部６ｂの先端に当接することにより、入力軸１８および弁作動部材１７の後退が停止し、入力軸１８および弁作動部材１７はともに後退限となる。しかしながら、入力軸１８および弁作動部材１７の後退が停止しても、パワーピストン８、反力ピストン２０、円錐弁１４および弁座部材１０は、ともに更に後退を続ける。このため、弁作動部材１７のストッパ部１７ｂが反力ピストン２０のストッパ部２０ｃから離隔するとともに、円錐弁１４が弁作動部材１７の第２弁座１７ａに近づいてくる。
【００７５】
パワーピストン８の右端がプラグ６に当接すると、パワーピストン８の後退が停止し、マスタシリンダピストン４７およびパワーピストン８は非作動位置となって、ブレーキが迅速にかつ完全に解除される。このブレーキ解除時では、ブレーキ液圧倍力装置１の出力は、入力の減少に対して、入力上昇時とは逆に、全負荷、直線γ、直線β、および直線アルファに沿って降下するようなる。
【００７６】
パワーピストン８の非作動位置では、円錐弁１４が弁作動部材１７の第２弁座１７ａにきわめて近づいて円錐弁１４と第２弁座１７ａとの間の間隙がきわめて小さくなり、着座寸前となる。したがってブレーキペダルが踏み込まれて入力軸１８および弁作動部材１７が前進すると、直ぐに第２弁座１７ａが円錐弁１４に着座するとともに円錐弁１４が弁座部材１０の第１弁座１０ａから直ぐに離座する。すなわち、制御弁６０の切換作動を行うためのロスストロークがきわめて小さくなり、ブレーキが迅速に作動する。
【００７７】
このようにして、ブレーキ操作時には迅速にブレーキが作動するとともに、ブレーキ操作解除時にはブレーキ作動が迅速に解除し、ブレーキ液圧発生装置１はきわめて応答性のよいものとなる。
この例のブレーキ液圧倍力装置１によれば、通常ブレーキ時のサーボ制御の途中において、動力室２５の液圧が所定圧、換言すればブレーキペダルからの入力が所定の大きさ以上の時、通常ブレーキ時のサーボ比より大きなサーボ比でサーボ制御を行う逆二段サーボ特性を発揮することができるようになる。その場合、入力軸１８の段部１８ａおよび反力ピストン２０の右端２０ｅが位置する場所に、反力室４１を設けるとともに、圧力切換弁４６により、この反力室４１に動力室２５の液圧を導入するという簡単な構造で、この逆二段サーボ特性を得ることができる。
【００７８】
これにより、急ブレーキ時には、ブレーキペダルをサーボ比切換点まで踏み込むだけで、従来のようにかなり大きく踏み込まなくても、早く大きなブレーキ力を得ることができる。また、運転に慣れていない運転者でも、急ブレーキ時には確実に大きなブレーキ力を発生させるように補助することができるようになる。
【００７９】
更に、圧力切換弁４６の設定作動圧を可変にして種々調整することにより、サーボ比切換点を変えることが可能となる。
【００８０】
なお、反力室４１に導入する液圧を、動力室２５の液圧に代えて、アキュムレータ４０のアキュムレータ圧を導入することもできる。この場合には、アキュムレータ圧を圧力調整弁で調圧して反力室４１に導入するようにする。アキュムレータ圧は、通常ブレーキ時には動力室２５の液圧より高いので、大きなサーボ比、すなわち同じ入力に対して大きな出力を得ることができる。このアキュムレータ圧導入については、後述する他の例で具体的に説明する。
【００８１】
図５は、本発明の実施の形態の第２例を示す、図１と同様の図である。なお、第１例と同じ構成要素には同じ符号を付すことにより、その詳細な説明は省略する。また、以下の各例の説明においても、その前の例と同じ構成要素には同様に同じ符号を付すが、符号を付さない構成要素も他の例の対応する構成要素と同じである。
【００８２】
前述の第１例では、サーボ比を変更するにあたって、動力室２５の液圧によって切換制御される圧力切換弁４６によって、反力室４１に動力室２５の液圧を導入するようにしているが、この第２例のブレーキ液圧倍力装置１では、圧力切換弁４６に代えて、同じ二位置三方弁からなる電磁切換弁６１を設けているとともに、出力口２７の液圧すなわち動力室２５の液圧を検出する圧力センサ６２を設けている。この圧力センサ６２の検出信号が図示しない電子制御装置に入力され、電子制御装置はこの検出信号に基づいて動力室２５の液圧が所定圧以上であると判断したとき、電磁切換弁６１を第２位置IIに切換設定するようになっている。
【００８３】
この第２例のブレーキ液圧倍力装置１においても、第１例と同じ図４に示す逆二段サーボ特性を有するようになる。
この第２例のブレーキ液圧倍力装置１の他の構成および作用効果は、前述の第１例と同じである。
【００８４】
図６は、本発明の実施の形態の第３例を示す、図１と同様の図である。
【００８５】
前述の第２例では、電磁切換弁６１により、反力室４１に対して動力室２５の液圧の給排を制御するようにしているが、この第３例のブレーキ液圧倍力装置１では、図６に示すように電磁切換弁６１に代えてリニアソレノイドバルブからなる電磁比例制御弁６３を設けているとともに、圧力調整弁６４を設けており、アキュムレータ４０のアキュムレータ圧を圧力調整弁６４によって調整し、この調整した液圧をこの電磁比例制御弁６３によって制御して反力室４１に導入するようにしている。その場合、電子制御装置は圧力センサ６２の検出信号に基づいて動力室２５の液圧が所定圧以上であることを判断したとき、圧力センサ６２の検出信号の大きさつまり動力室２５の液圧に比例した大きさの制御信号を出力して電磁比例制御弁６３を作動するようになっている。そして、電磁比例制御弁６１は、通常時は反力室４１をリザーバ３３に接続しているが、作動時は反力室４１をリザーバ３３から遮断し、この反力室４１に、アキュムレータ圧を調圧した液圧を電子制御装置からの制御信号の大きさに比例させて制御した液圧を導入するようになっている。
この第３例のブレーキ液圧倍力装置１の他の構成は、前述の第１および第２例と同じである。
【００８６】
このように構成された第３例のブレーキ液圧倍力装置１においては、圧力センサ６２からの検出信号に基づいて、電子制御装置は、動力室２５の液圧が電磁比例制御弁６３の設定作動圧より低いと判断したときは、電磁比例制御弁６３を作動しなく、１電磁比例制御弁６１は反力室４１をリザーバ３３に接続する。したがって、このときは第１および第２例とまったく同じであり、ブレーキ液圧倍力装置１の入出力特性は、直線αに沿うサーボ特性つまりジャンピング特性および直線βに沿う通常ブレーキ時のサーボ特性となる。動力室２５の液圧が電磁比例制御弁６３の設定作動圧になったことが判断されると、電磁比例制御弁６３が作動され、この電磁比例制御弁６３は作動して反力室４１をリザーバ３３から遮断するとともに、アキュムレータ圧を調圧した液圧を動力室２５の液圧に比例した液圧を反力室４１に導入する。これにより、前述の第１および第２例と同様にサーボ比が大きい方に変化する。反力室４１に導入された液圧は、動力室２５の液圧に比例して上昇するようになるので、サーボ比も比例して次第に大きくなる。このときの、ブレーキ液圧倍力装置１の入出力特性は、
【００８７】
【数４】

【００８８】
で与えられ、図７に示す直線δに沿うサーボ特性となる。ここで、Ｐ₂は反力室４１の液圧である。
【００８９】
反力室４１の液圧Ｐ₂が動力室２５の液圧Ｐ₁に等しくなると、前述の第１例とまったく同じになり、ブレーキ液圧倍力装置１は、直線γに沿う大きなサーボ比によるサーボ制御を行い、そのときのブレーキ液圧倍力装置１の入出力特性は、数式２で与えられる。
【００９０】
この第３例のブレーキ液圧倍力装置１によれば、直線βのサーボ比から直線γのサーボ比への切換えを、直線δに沿って滑らかに行うことができるようになる。また、反力室４１にアキュムレータ４０の蓄圧を導入するようにしているので、反力室４１に動力室２５の液圧を導入する場合に比べて、ブレーキ液圧倍力装置１の出力を大きくできる。
第３例のブレーキ液圧倍力装置１の他の作用効果は、第１および第２例と同じである。
【００９１】
なお、反力室４１の液圧を制御することにより、図７に二点鎖線で示すように直線βのサーボ比から直接全負荷に滑らかに変えるようにすることもできる。また、この第３例においても、アキュムレータ圧に代えて、反力室４１に動力室２５の液圧を電磁比例制御弁６３で制御して導入することもできる。
【００９２】
図８は、本発明の実施の形態の第４例を示す、図２と同様の図である。
前述の第１ないし第３例の各ブレーキ液圧倍力装置１では、反力ピストン２０を付勢するスプリング２１を設けてジャンピング特性を有するようにしているが、この第４例のブレーキ液圧倍力装置１では、反力ピストン２０を付勢するスプリング２１を削除して反力ピストン２０をフリーにし、ブレーキ作動初期のジャンピング特性がないようにしている。
この第４例のブレーキ液圧倍力装置１の他の構成は、第１ないし第３例のいずれかと同じである。
【００９３】
このように構成された第４例のブレーキ液圧倍力装置１においては、ブレーキ操作時、入力軸１８が前進して動力室に液圧が導入されると、この動力室の液圧により、直ぐに反力ピストン２０が入力軸１８に対して後方へ相対移動し、反力ピストン２０の右端２０ｅが入力軸１８の段部１８ａに直ぐに当接する。このため、入力軸１８のパワーピストン８側の外径が反力ピストン２０の外径により入力軸１８のパワーピストン８側の外径より大きくなった状態となる。また、第３例のブレーキ液圧倍力装置１は、直ぐに通常ブレーキ時のサーボ比によるサーボ制御を開始するようになる。このときのブレーキ液圧倍力装置１の入出力特性は、
【００９４】
【数５】

【００９５】
で与えられるとともに、図９に示すように直線β′で表される。すなわち、第１および第２例のような制動初期のジャンピング特性が行われない。
【００９６】
第１例の場合と同様に、動力室２５の液圧が圧力切換弁４６の設定作動圧になると、反力室４１には動力室２５の液圧が導入される。したがって、ブレーキ液圧倍力装置１のサーボ比が通常ブレーキ時のサーボ比より大きなサーボ比に変更される。このときのブレーキ液圧倍力装置１の入出力特性は、
【００９７】
【数６】

【００９８】
で与えられるとともに、図９に示すように直線γ′で表される。
この第４例のブレーキ液圧倍力装置１の他の作用効果は、第１ないし第３例のいずれかと同じである。
【００９９】
なお、この第４例の場合は、反力ピストン２０は必ずしも必要ではなく、省略できる。この場合は、入力軸１８の動力室２５に面する部分（すなわち弁作動部材１７のフランジ部１７ｂを含む部分）の断面積が、入力軸８のブレーキペダル側部分の断面積より大きく、かつ入力軸１８の、反力室４１に面する部分に、反力室４１の液圧が入力軸１８に対して入力と同方向に作用する受圧部、つまり段付の入力軸１８のパワーピストン８側の外径を入力軸１８のブレーキペダル側の外径より大きく形成することが必要である。
【０１００】
また、この第４例のブレーキ液圧倍力装置１において、前述の第２例のように圧力切換弁４６に代えて電磁切換弁６１を用いることができる。更に、第３例のように圧力センサ６２および電磁比例制御弁６３を用いて、反力室４１に、アキュムレータ圧を調圧した液圧を動力室２５の液圧に応じて制御して導入することにより、サーボ比を滑らかに変更するようにすることもできる。この場合のブレーキ液圧倍力装置１の入出力特性は、
【０１０１】
【数７】

【０１０２】
で与えられ、図１０に示す示すように直線β′、δ′、γ′で表される。この場合のブレーキ液圧倍力装置１の他の作用効果は、第３例と同じである。
【０１０３】
図１１は、本発明の実施の形態の第５例を示す、図１と同様の図である。
この第５例のブレーキ液圧倍力装置１は、前述の第４例と同様に、スプリング２１が削除されていて、反力ピストン２０によるジャンピング特性を有さないものとされている。
【０１０４】
また、この第５例のブレーキ液圧倍力装置１は、マスタシリンダ２が、プライマリピストン４７′とセカンダリピストン４７″とを有するタンデムマスタシリンダ２として構成されている。そして、プライマリピストン４７′がパワーピストン８の前端に一体に設けられている。また、セカンダリピストン４７″はその後端を除いて前述の各例のマスタシリンダピストン４７と同じに構成されている。
【０１０５】
両ピストン４７′,４７″の間隔を規制する間隔規制ロッド６５がマスタシリンダピストン４７の方へ突出して固定されているとともに、この間隔規制ロッド６５にリテーナ６６が軸方向に摺動可能に嵌合されている。また、リテーナ６６とプライマリピストン４７′の前端との間には、スプリング６７が縮設されていて、リテーナ６６が常時プライマリピストン４７′から遠ざかる方向に付勢されている。通常時は、このリテーナ６６は間隔規制ロッド６５の頭部６５ａに当接して、それ以上プライマリピストン４７′から離れることを規制されている。
【０１０６】
径方向孔３１がセカンダリピストン４７″に設けられていることから、制御弁６０の弁体１５の軸方向孔３０は、パワーピストン８に形成された通路孔６８、段付孔４の小径部４ａ、およびハウジング３に穿設された通路孔６９を介してリザーバ３３に常時接続されている。
【０１０７】
セカンダリピストン４７″の後端は、リターンスプリング５９のばね力により、通常時はリテーナ６６に当接されている。プライマリピストン４７′の前端とセカンダリピストン４７″の後端には、それぞれカップシール７０,７１が設けられており、これらのカップシール７０,７１の間の小径部４ａに、液室７２が画成されている。両カップシール７０,７１は、それぞれカップシール７０,７１を通る液室７２外から液室７２内への液の流れは許容するが、カップシール７０,７１を通る液室７２から液室７２外への液の流れは阻止するようになっている。
【０１０８】
そして、パワーピストン８の有効受圧面積、プライマリピストン４７′の有効受圧面積、およびセカンダリピストン４７″の前後端部の各有効受圧面積はすべて等しく設定されている。
【０１０９】
制御圧導入口４５には、可変サーボ装置７３が接続されている。この可変サーボ装置７３は、ブレーキ液圧倍力装置１にジャンピング特性を持たせるための第１切換弁７４と、この第１切換弁７４に直列に配設されて、ブレーキ液圧倍力装置１に逆二段サーボ特性を持たせるための第２切換弁７５とから構成されている。
【０１１０】
第１切換弁７４は二位置三方弁からなり、動力室２５の液圧によるパイロット圧で制御されるようにされている。そして、この第１切換弁７４は、第２切換弁７５を出力口２７に接続する第１位置Iと、第２切換弁７５をリザーバ３３に接続する第２位置IIとが設定されており、ブレーキ液圧倍力装置１の非作動時は第１位置Iに設定されているとともに、動力室２５の液圧が、ジャンピング作用を終了させる第１設定作動圧（反力ピストン２０を付勢するスプリング２１を有する前述の例における、反力ピストン２０の作動圧に相当する）となると、第２位置IIに切り換え設定されるようになっている。
【０１１１】
また、第２切換弁７５も同様に二位置三方弁からなり、動力室２５の液圧によるパイロット圧で制御されるようにされている。そして、この第２切換弁７５は、
制御圧導入口４５を第１切換弁７４に接続する第１位置Iと、制御圧導入口４５を出力口２７に接続する第２位置IIとが設定されており、通常時は第１位置Iに設定されているとともに、動力室２５の液圧が、サーボ比切換点の液圧に設定された第２設定作動圧になると第２位置IIに切り換え設定されるようになっている。
【０１１２】
ところで、この第５例のブレーキ液圧倍力装置１は、サーボ比を変更する可変サーボ装置７３の他にもいくつかの装置を備えている。
すなわち、液室７２は、ハウジング３の接続口７６を介して可変ストローク装置７７に接続されている。この可変ストローク装置７７は、ピストン７８、シリンダ７９、およびスプリング８０を有し、ペダルストロークを確保するためのストロークシュミレータ８１と、接続口７６とストロークシュミレータ８１との間の通路に設けられたオリフィス８２と、このオリフィス８２をバイパスして設けられ、ストロークシュミレータ８１から接続口７６への液の流れのみを許容するチェックバルブ８３とから構成されている。
【０１１３】
また、ハウジング３に接続口８４が穿設されており、この接続口８４は、液圧回路３６から分岐された液通路８５に接続されている。液通路８５には、自動ブレーキ装置８６が設けられており、この自動ブレーキ装置８６は、二位置三方弁からなる電磁切換弁８７と、圧力調整弁８８とから構成されている。この電磁切換弁８７は、接続口８４をリザーバ３３に接続する第１位置Iと、接続口７９を、圧力調整弁８８を介してアキュムレータ４０に接続する第２位置IIとが設定されており、通常時は第１位置Iに設定されているとともに、自動ブレーキ作動時に第２位置IIに切り換え設定されるようになっている。そして、プライマリピストン４７′の非作動位置では、カップシール７０が通路孔６９と接続口８４との間に位置するようにされている。したがって、非作動時には、液は液室７２と接続口８４との間で両方向に自由に流れるようになっているが、パワーピストン８が前進して、カップシール７０が接続口８４を通り過ぎると、接続口８４から液室７２に向かう液の流れは許容されるが、液室７２から接続口８４に向かう液の流れは阻止されるようになっている。
【０１１４】
ブレーキ液圧倍力装置１の出力口２７と一方の系統のホイールシリンダ２８,２９とを接続する液通路には、ポンプ３８およびアキュムレータ４０の液圧源の失陥時に、一方の系統のブレーキを確実に作動するための液圧失陥時ブレーキ作動装置８９が設けられている。この液圧失陥時ブレーキ作動装置８９は、ピストン９０、シリンダ９１、およびスプリング９２を有する圧力変換シリンダ９３と、二位置三方弁からなる圧力制御切換弁９４とから構成されている。
【０１１５】
圧力変換シリンダ９３は、ポンプ３８から吐出された圧液が導入されたとき、ピストン９０が作動してブレーキ液圧を発生し、このブレーキ液圧はホイールシリンダ２８,２９に導入されるようになっている。また、圧力変換シリンダ９３は、液圧失陥がホイールシリンダ２８,２９側の失陥によるものである場合、ポンプ３８から吐出された液（つまり、アキュムレータ蓄圧液）がこの失陥部から外へ漏出するのを防止するようになっている。
【０１１６】
圧力制御切換弁９４は、アキュムレータ４０の蓄圧によるパイロット圧で制御されるようにされている。また、この圧力制御切換弁９４は、圧力変換シリンダ９３を出力口２７に接続する第１位置Iと、圧力変換シリンダ９３を接続口７６に接続する第２位置IIとが設定されており、液圧正常時は第１位置Iに設定されているとともに、液圧失陥時は第２位置IIに切り換え設定されるようになっている。 この第５例のブレーキ液圧倍力装置１の他の構成は、第４例と同じである。
【０１１７】
このように構成された第５例のブレーキ液圧倍力装置１においては、液圧正常時でかつ非作動時は、ブレーキ液圧倍力装置１は図示の状態になっている。この状態で、通常ブレーキ操作が行われ、動力室２５に液圧が導入されると、パワーピストン８が前進するとともに、反力ピストン２０は直ぐに作動して、その右端２０ｅが入力軸１８の段部１８ａに当接すると同時に、動力室２５の液圧が直ぐに反力室４１に導入される。これにより、ブレーキ液圧倍力装置１は、前述の例と同様に通常ブレーキ時のサーボ比より大きなサーボ比でサーボ制御を行うようになる。このときの、ブレーキ液圧倍力装置１の入出力特性は、前述の数式６で与えられるとともに、図１２に示すように直線α″で表される。したがって、この直線α″のサーボ比によるサーボ制御は、反力ピストン２０によるジャンピング作用とほぼ同じジャンピング作用を行うようになる。
【０１１８】
パワーピストン８の前進でプライマリピストン４７′も前進し、そのカップシール７０が接続口８４を通過すると、液室７２がリザーバ３３から遮断されて密封状態となる。このため、プライマリピストン４７′の前進で、液室７２の液が接続口７６から、可変ストローク装置７７のストロークシュミレータ８１に送られる。このとき、プライマリピストン４７′は、通常ブレーキ作動時であるため通常の速度で前進するので、ストロークシュミレータへ流動する液に対するオリフィス８２のオリフィス効果は小さい。したがって、プライマリピストン４７′すなわちパワーピストン８は通常の速度で、ピストン７８のストローク分つまりストロークシュミレータ８１のストロークシュミレータ分、ストロークするようになる。
【０１１９】
動力室２５の液圧が第１設定作動圧になると、第１切換弁７４が作動して第２位置IIに切換設定され、反力室４１が出力口２７から遮断されるとともにリザーバ３３に接続される。このため、反力室４１の液圧はリザーバ３３に排出され、前述の例と同様にサーボ比が小さくなって通常ブレーキ時のサーボ比と同じになり、これ以後、ブレーキ液圧倍力装置１は、通常ブレーキ時のサーボ比でサーボ制御を行うようになる。このときの、ブレーキ液圧倍力装置１の入出力特性は、前述の数式５で与えられるとともに、図１２に示すように直線β″で表される。
【０１２０】
動力室２５の液圧が第２設定作動圧になると、第２切換弁７５が作動して第２位置IIに切換設定され、反力室４１がリザーバ３３から遮断されるとともに再び出力口２７に接続される。このため、反力室４１には、動力室２５の液圧が再び導入され、前述の例と同様にサーボ比が通常ブレーキ時のサーボ比より大きいサーボ比でのサーボ制御を行うようになる。このときの、ブレーキ液圧倍力装置１の入出力特性は、前述の数式６で与えられるとともに、図１２に示すように直線α″と同じ直線である直線γ″で表される。
【０１２１】
このように、この第５例のブレーキ液圧倍力装置１においては、ブレーキ作動初期段階で、直線α″に沿った大きなサーボ比のサーボ制御によるジャンピング作用が行われ、その後直線β″に沿った小さなサーボ比のサーボ制御による通常ブレーキ作用が行われ、入力がある程度大きくなると、再び直線γ″に沿った大きなサーボ比のサーボ制御による、急ブレーキ等の大きなブレーキ力を必要とするブレーキ制御が行われるようになる。
【０１２２】
こうして、第５例のブレーキ液圧倍力装置１は反力ピストン２０によるジャンピング特性を有さなくても、簡単な構造で、第１切換弁７４によってジャンピング特性を有するようになるとともに、第２切換弁７５によって逆二段サーボ特性を有するようになる。
【０１２３】
ところで、この第５例のブレーキ液圧倍力装置１は、ジャンピング作用および逆二段サーボ特性による可変サーボ作用を行うばかりでなく、可変ストローク作用、自動ブレーキ作用、および液圧失陥時のブレーキ作用も行うようになっている。
【０１２４】
まず、可変ストロークの作用について説明する。運転者が通常ブレーキ時の速度でブレーキペダルを踏み込んだ場合は、前述の通りである。また、ブレーキペダルを急速に踏み込んで急ブレーキをかけると、パワーピストン８およびプライマリピストン４７′も急速に前進するので、液室７２の液は急速に接続口７６からストロークシュミレータ８１に送給される。このとき、接続口７６からの液の流速が速いので、オリフィス８２によるオリフィス効果が大きくなり、液室７２に高い液圧が発生し、この液室７２の高い液圧がプライマリピストン４７′およびパワーピストン８を介して入力軸１８に大きな反力として作用するので、ペダルストロークは通常ブレーキ時により小さくなる。したがって、この大きな反力により、液圧倍力装置１は大きな出力を発生し、この大きな出力によりマスタシリンダ２のセカンダリピストン４７″が高いマスタシリンダ圧を発生するとともに、このセカンダリピストン４７″の有効受圧面積とパワーピストン８の有効受圧面積とが同じであることから、動力室２５の液圧がこのマスタシリンダ圧と等しい高い液圧となる。
【０１２５】
そして、動力室２５の高い液圧が出力口２７から圧力変換シリンダ９３に供給され、圧力変換シリンダ９３のピストン９０が作動して高圧のブレーキ液圧を発生し、この高いブレーキ液圧がホイールシリンダ２８,２９に導入され、ホイールシリンダ２８,２９は大きなブレーキ力を発生する。一方、高いマスタシリンダ圧が出力口５０からホイールシリンダ５１,５２に導入され、ホイールシリンダ５１,５２は大きなブレーキ力を発生する。
【０１２６】
ブレーキペダルの解放によりブレーキ作動が解除されると、ストロークシュミレータ８１に供給された液は、チェックバルブ８３により遅れることなく液室７２の方へ戻されるので、オリフィス８２が設けられても、パワーピストン８および入力軸１８は遅れることなく、非作動位置に戻るようになる。
【０１２７】
このようにして、オリフィス８２によりブレーキペダルの踏み込み速度に応じてペダルストロークを変えることができ、ブレーキペダルの急速な踏み込み時には、可変ストローク装置７７により、液圧倍力装置１は入力軸１８の小さいストロークで大きな出力を発生するので、ブレーキ力の立ち上がりが早くなるとともに、両系統に大きなブレーキ力が迅速に発生されるようになる。
【０１２８】
次に、自動ブレーキ作用について説明する。車両走行中に、自動ブレーキ作動条件が成立すると、図示しない電子制御装置が電磁切換弁８７を第２位置IIに切換設定される。このため、接続口８４が圧力調整弁８８を介してアキュムレータ４０に接続される。すると、アキュムレータ４０の蓄圧が圧力調整弁８８によって所定圧に調整され、この調整された液圧が接続口８４を通って液室７２に導入されるとともに、この液圧により圧力制御切換弁９４が第２位置IIに切換設定される。液室７２に導入された液圧は、更に接続口７６を通って可変ストローク装置７７に導入されるとともに、液圧失陥時ブレーキ作動装置８９の圧力変換シリンダ９３に導入される。すると、ピストン９０が作動してブレーキ液圧を発生し、発生したブレーキ液圧がホイールシリンダ２８,２９に導入されて、一方の系統のブレーキが作動する。
【０１２９】
一方、液室７２に導入されかつ調整された液圧は、マスタシリンダ２のセカンダリピストン４７″の後端面に作用するようになるので、セカンダリピストン４７″が作動し、液圧４９に、液室７２の液圧と等しいマスタシリンダ圧を発生する。このマスタシリンダ圧がホイールシリンダ５１,５２に導入されて、他方の系統のブレーキが作動する。こうして、自動ブレーキが両系統に確実に作動するようになる。
【０１３０】
自動ブレーキの作動解除の条件が成立すると、電子制御装置は電磁切換弁８７を再び非作動の第１位置Iに設定する。このため、接続口７６がリザーバ３３に接続されるので、液室７２および圧力変換シリンダ９３の液圧はリザーバ３３に排出される。このとき、接続口８４の液圧がある程度低下すると、圧力制御切換弁９４が第１位置Iに復帰されて、圧力変換シリンダ９３が動力室２５に接続されるので、圧力変換シリンダ９３の圧液は動力室２５を通ってリザーバ３３に排出される。これにより、圧力変換シリンダ９３のブレーキ液圧が消滅し、一方の系統のブレーキが解除される。
【０１３１】
また、液室７２がリザーバ３３に接続されてその液圧が低下することにより、セカンダリピストン４７″も後退し、セカンダリピストン４７″が非作動位置に戻ったときは、前述の通常ブレーキの場合と同様に液室４９がリザーバ３３に連通するので、マスタシリンダ圧が消滅し、他方の系統のブレーキが解除される。こうして、自動ブレーキが完全に解除される。
【０１３２】
次に、液圧失陥時のブレーキ作用について説明する。ポンプ３８およびアキュムレータ４０等の液圧源の液圧が失陥すると、圧力制御切換弁９４が第２位置IIに設定される。この状態で、運転者がブレーキペダルの踏み込みによる通常ブレーキ操作を行って入力軸１８を前進させ、制御弁６０を切り換えても、動力室２５には液圧が導入されない。このため、パワーピストン８は、動力室２５の液圧によっては作動しない。更に、ブレーキペダルが大きく踏み込まれて入力軸１８が大きく前進すると、弁体１５が最大ストロークしてパワーピストン８に当接して、このパワーピストン８を押すようになる。すると、パワーピストン８と一体のプライマリピストン４７′が前進し、そのカップシール７０が接続口８４を通過すると、液室７２に液圧が発生し、この液圧が接続口７６を介して圧力変換シリンダ９３に導入される。これ以後の一方の系統のブレーキは、前述の自動ブレーキの場合と同じようにして作動する。
【０１３３】
更に、この液圧失陥時のブレーキ作動においては、プライマリピストン４７′が前進することにより、セカンダリピストン４７″も前進し、前述と同様に弁５４が弁座５８に着座して、液室４９に液圧が発生する。液室４９の液圧は出力口５０を介して他方の系統のホイールシリンダ５１,５２に導入され、他方の系統のブレーキも作動する。このとき、セカンダリピストン４７″の前後端の各有効受圧面積が等しいので、液室７２の液圧と液室４９の液圧は同じになり、その結果、両系統のブレーキ力は同じになる。
【０１３４】
液圧失陥時におけるブレーキ作動の解除は、通常ブレーキの解除と同様にブレーキペダルを解放することにより行われる。ブレーキペダルの解放により、パワーピストン８とともにプライマリピストン４７′が後退して、液室７２の液圧が低下するので、一方の系統のブレーキ力が低下するとともに、セカンダリピストン４７″が後退して、液室４９の液圧が低下するので、他方の系統のブレーキ力も低下する。更に、プライマリピストン４７′が後退して、カップシール７０が接続口８４を通過すると、液室７２が接続口８４に連通する。すると、液室７２がリザーバ３３に連通するので、液室７２および圧力変換シリンダ９３の液圧がリザーバ３３に排出されるので、一方の系統のブレーキが完全に解除される。また、マスタシリンダピストン４７も更に後退するので、通常ブレーキの作動解除と同様に、弁５４が弁座５８から離座するので、液室４９がリザーバ３３に連通し、他方の系統のブレーキも完全に解除される。
この第５例のブレーキ液圧倍力装置１の他の作用効果は、第２例と同じである。
【０１３５】
図１３は、本発明の実施の形態の第６例を示す、図１１と同様の図である。 この第６例のブレーキ液圧倍力装置１は、前述の図１１に示す第５例の可変サーボ装置７３における動力室２５の液圧により制御される第１および第２切換弁７４,７５に代えて、二位置三方弁からなる電磁切換弁９５が設けられている。この電磁切換弁９５は、非作動時に制御圧導入口４５を出力口２７に接続する第１位置Iと、作動時に制御圧導入口４５をリザーバ３３に接続する第２位置IIとが設定されている。また、圧力センサ６２への液圧導入通路に、オリフィス１００と、このオリフィス１００をバイパスして設けられ、圧力センサ６２から排出される方向の液の流れのみを許容するチェックバルブ１０１とを備えている。
【０１３６】
そして、電子制御装置は、圧力センサ６２によって検出された出力口の液圧つまりは動力室２５の液圧が従来のジャンピング作用を終了するジャンピング終了圧になるまでは、この電磁切換弁９５を非作動の第１位置Iに設定し、また動力室２５の液圧がジャンピング作用終了圧になったとき、電磁切換弁９５を作動して第２位置IIに設定し、更に動力室２５の液圧がサーボ比切換点の圧力となったとき、電磁切換弁９５を再び非作動の第１位置Iに設定するようになっている。
【０１３７】
可変ストローク装置７２は、第１電磁開閉弁１０２を介して接続口７６に接続されている。この第１電磁開閉弁１０２は、連通位置Iと遮断位置IIとが設定されており、通常時は連通位置Iに設定される常開弁とされている。更に、可変ストローク装置７２は、接続口７６と圧力変換シリンダ９３とを接続する通路に設けられた第２電磁開閉弁１０３を備えている。この第２電磁開閉弁１０３は、遮断位置Iと連通位置IIとが設定されており、通常時は遮断位置Iに設定される常閉弁とされている。
【０１３８】
前述の第５例の液圧失陥時ブレーキ作動装置８９における圧力制御切換弁９４は、アキュムレータ４０の蓄圧によって切換制御されるようになっているが、この第６例の液圧失陥時ブレーキ作動装置８９においては、この圧力制御切換弁９４は、二位置三方弁からなる電磁切換弁で構成されている。そして、この電磁切換弁の切換制御のために、アキュムレータ４０の蓄圧を検出する圧力センサ１０４が設けられている。
【０１３９】
そして、電子制御装置は、自動ブレーキ作動条件が成立したと判断したとき、および圧力センサ１０４からの検出信号に基づいてアキュムレータ４０の蓄圧が失陥したと判断したときに、圧力制御切換弁９４を切換設定するようになっている。
この第６例のブレーキ液圧倍力装置の他の構成は、第５例と同じである。
【０１４０】
このように構成されたこの第６例のブレーキ液圧倍力装置の作用について説明する。
ブレーキ非作動時は、ブレーキ液圧倍力装置１およびマスタシリンダ２の各構成要素は、図１３に示す非作動位置にある。この状態で、ブレーキペダルの通常速度の踏み込みで通常ブレーキ操作が行われると、前述のように動力室２５に圧液が導入されるとともに、この動力室２５の圧液は、更に出力口２７、電磁切換弁９５を通って制御圧導入口４５から反力室４１に導入される。このとき、通常速度でブレーキペダルが踏み込まれるので、液はオリフィス１００によってほとんど絞られることなく、反力室４１に導入される。こうして、前述の第５例と同様に、ブレーキ液圧倍力装置１は、図１２に示す直線α″のサーボ比によるサーボ制御、すなわちジャンピング作用を行うようになる。
【０１４１】
動力室２５の液圧がジャンピング作用を終了する大きさになると、圧力センサ６２の検出信号に基づいて、電子制御装置が電磁切換弁９５を第２位置IIに切り換える。すると、反力室４１はリザーバ３３に接続され、反力室４１に導入された圧液はチェックバルブ１０１および電磁切換弁９５を通ってリザーバ３３に遅れることなく排出される。これにより、ブレーキ液圧倍力装置１は、図１２に示す直線β″の小さいサーボ比による通常のサーボ制御を行うようになる。動力室２５の液圧がサーボ比切換点になると、電磁切換弁９５が再び第１位置Iに切り換えられる。すると、反力室４１は再び出力口２７に接続されるので、動力室２５の圧液が反力室４１に導入され、反力室４１の液圧が動力室２５の液圧と同じになる。これにより、ブレーキ液圧倍力装置１は、図１２に示す直線γ″の大きいサーボ比によるサーボ制御を行うようになる。
【０１４２】
ところで、ブレーキペダルが通常速度よりはるかに急速に踏み込まれて急ブレーキ操作が行われると、動力室２５の液圧が急速に上昇する。すると、動力室２５の圧液が反力室４１へ急速に流動するようになるが、オリフィス１００により圧力センサ６２の液圧上昇が遅れるので、電磁切換弁９５の切換が遅れる。この遅れの分、図１２に示す直線α″の大きなサーボ比から直線β″の小さなサーボ比への移行が遅くなる。つまり、サーボ比切換点が変わり、ジャンピング作用の終了が遅くなる。したがって、その分ブレーキ液圧倍力装置１は大きな出力を発生するようになり、急ブレーキ時により大きなブレーキ力を得ることができる。
【０１４３】
一方、可変ストローク装置７７においては、各電磁弁９４,１０２,１０３が図１３に示す非作動位置にあるときは、前述の第５例とまったく同じように作動し、ブレーキペダルは、ストロークシュミレータ８１のストロークシュミレータ分だけストロークする。また、第１電磁開閉弁１０２のみが作動して遮断位置IIに設定されると、液室７２がストロークシュミレータ８１から遮断されてロック状態となり、ペダルストロークはマスタシリンダ２のセカンダリピストン４７″側の他方の系統のストローク分のみとなる。更に、第１および第２電磁開閉弁１０２,１０３と電磁切換弁９４とが作動してともに位置IIに設定されると、圧力変換シリンダ９３が動力室２５から遮断されかつ液室７２に接続されるとともに、液室７２がストロークシュミレータ８１から遮断される。このときは、ペダルストロークはホイールシリンダ２８,２９のストローク分となる。更に、第２電磁開閉弁１０３と電磁切換弁９４とが作動してともに位置IIに設定されると、圧力変換シリンダ９３が動力室２５から遮断されかつ液室７２に接続される。このときは、液室７２はストロークシュミレータ８１に接続されるので、ペダルストロークはストロークシュミレータ８１のストロークシュミレータ分とホイールシリンダ２８,２９のストローク分とを加えたものとなる。
【０１４４】
このように、各電磁弁９４,１０２,１０３の作動を制御することにより、ペダルストロークを種々変えることができるようになる。したがって、積載状態等の車両状況、ブレーキ状況、あるいは運転者等によって、より適正なペダルストロークを設定することができる。
【０１４５】
また、自動ブレーキ装置８６においては、自動ブレーキ作動条件が成立すると、電子制御装置が、電磁切換弁８７を第２位置IIに切り換えるだけでなく、電磁切換弁９４も第２位置IIに切り換える。したがって、前述の第５例の場合とまったく同じになり、自動ブレーキが作動する。
【０１４６】
更に、液圧失陥時ブレーキ作動装置８９においては、アキュムレータ４０の液圧が失陥すると、圧力センサ１０４からの検出信号により、電子制御装置は電磁切換弁９４を第２位置IIに切り換える。したがって、前述の第５例の場合とまったく同じになり、液圧失陥時でも、ブレーキペダルの踏み込みによりブレーキを作動させることが可能となる。
【０１４７】
図１４は、本発明の実施の形態の第７例を部分的に示す図である。
前述の図１１に示す第５例では、第１および第２切換弁７４,７５を用いて、ジャンピング特性と逆二段サーボ特性とを得るようにしているが、この第７例のブレーキ液圧倍力装置１は、図１４に示すように第３および第４電磁開閉弁１０５,１０６が用いられている。
【０１４８】
第３電磁開閉弁１０５は、制御圧導入口４５とリザーバ３３とを接続する通路に設けられて、制御圧導入口４５とリザーバ３３とを遮断する遮断位置Iと、制御圧導入口４５とリザーバ３３とを連通する連通位置IIとが設定されており、通常は遮断位置Iに設定される常閉の開閉弁とされている。
【０１４９】
第４電磁開閉弁１０６は、制御圧導入口４５と出力口２７とを接続する通路に設けられて、制御圧導入口４５と出力口２７とを連通する連通位置Iと、制御圧導入口４５と出力口２７とを遮断する遮断位置IIとが設定されており、通常は連通位置IIに設定される常開の開閉弁とされている。更に、これらの第３および第４電磁開閉弁１０５,１０６の開閉は、動力室２５の液圧によって制御されるようになっており、そのために前述と同様の、動力室２５の液圧を検出する圧力センサ６２が設けられている。
この第７例のブレーキ液圧倍力装置１の他の構成は、第５例と同じである。
【０１５０】
このように構成された第７例のブレーキ液圧倍力装置１においては、図１４に示す非作動状態から、通常ブレーキ操作が行われると、第５例とまったく同様にして、ブレーキ液圧倍力装置１は図１２に示す直線α″のサーボ比によるジャンピング作用を行う。動力室２５の液圧がジャンピング作用の終了に対応する圧力になると、圧力センサ６２からの検出信号に基づいて、電子制御装置が第３電磁開閉弁１０５を連通位置Iに、また第４電磁開閉弁１０６を遮断位置IIに設定する。すると、反力室４１がリザーバ３３に接続されるので、ブレーキ液圧倍力装置１は図１２に示す直線β″のサーボ比による通常ブレーキのサーボ制御を行う。動力室２５の液圧がサーボ比切換点の圧力になると、圧力センサ６２からの検出信号に基づいて、電子制御装置が第３電磁開閉弁１０５を遮断位置Iに、また第４電磁開閉弁１０６を連通位置Iに設定する。すると、反力室４１が出力口２７に接続されるので、ブレーキ液圧倍力装置１は図１２に示す直線γ″の大きなサーボ比によるサーボ制御を行う。
この第７例のブレーキ液圧倍力装置１の他の構成および作用効果は、第５例と同じである。
【０１５１】
なお、図１１に示す第５例のブレーキ液圧倍力装置１の第１切換弁７４へのパイロット圧の導入通路に、第６および第７例と同様のオリフィス１００とチェックバルブ１０１とを設けて、急ブレーキ時に、第１切換弁７４の第２位置IIへの切換を遅らせることにより、同様にしてより大きなブレーキ力を得るようにすることもできる。
【０１５２】
図１５は、本発明の実施の形態の第８例を部分的に示す図である。
図１５に示すように、この第８例のブレーキ制御装置１の可変サーボ装置７３では、図６に示す電磁比例制御弁６３および圧力調整弁６４が設けられており、アキュムレータ４０のアキュムレータ圧が圧力調整弁６４によって調整され、更に電磁比例制御弁６３によって制御されて反力室４１に導入されるようになっている。
この第８例のブレーキ液圧倍力装置１の他の構成は、第５例と同じである。
【０１５３】
この第８例のブレーキ液圧倍力装置１においては、非作動時は、電磁比例制御弁６３が反力室４１をリザーバ３３から遮断されかつ最大開弁量で圧力調整弁６４に接続しており、したがってアキュムレータ４０の蓄圧を圧力調整弁６４によって調圧した液圧が反力室３３に導入されている。この状態で、通常ブレーキ操作が行われると、ブレーキ液圧倍力装置１は、図１２に示す直線α″の大きなサーボ比によるサーボ制御、すなわちジャンピング作用を行うようになる。
【０１５４】
動力室２５の液圧がジャンピング作用を終了する大きさになると、圧力センサ６２の検出信号に基づいて、電子制御装置は電磁比例制御弁６３を作動制御するので、電磁比例制御弁６３は反力室４１の液圧を動力室２５の液圧に応じた圧力に制御する。これにより、ブレーキ液圧倍力装置１は、図１２に示す直線β″の小さいサーボ比による通常のサーボ制御を行うようになる。動力室２５の液圧がサーボ比切換点になると、電磁比例制御弁６３が再び非作動となり、反力室４１に圧力調整弁６４によって調圧した液圧が導入される。これにより、ブレーキ液圧倍力装置１は、再び図１２に示す直線α″と同じ直線γ″の大きいサーボ比によるサーボ制御を行うようになる。
このように、この第８例のブレーキ液圧倍力装置１においても、ジャンピング作用と逆二段サーボ作用とを行うようになる。
この第８例のブレーキ液圧倍力装置１の他の作用効果は、第５例と同じである。
【０１５５】
なお、小さいサーボ比から大きいサーボ比へ移行する際に、電磁比例制御弁６３を適宜制御することにより、前述の図７および図１０に直線δ,δ′で示すと同様に滑らかに移行することができる。また、反力室４１に導入される液圧として、アキュムレータの蓄圧に代えて動力室２５の液圧を用いることもできる。
【０１５６】
ところで、前述の第５ないし第８例では、パワーピストン８側が小径でブレーキペダル側が大径の段付の入力軸１８と反力ピストン２０とを用いて、入力軸１８のパワーピストン８側をブレーキペダル側より大径にする段部を形成し、この段部を反力室４１に位置するようにしているが、以後の各例では反力ピストン２０を省略し、パワーピストン８側が小径でブレーキペダル側が大径の段付の入力軸１８の段部を反力室４１に位置するようにしている。
【０１５７】
図１６は、本発明の実施の形態の第９例を部分的に示す図である。
図１１に示す第５例では、反力ピストン２０により、入力軸１８のパワーピストン８側の外径が実質的に反力ピストン２０の外径（受圧面積Ａ₃）となるため、ブレーキペダル側の外径（受圧面積Ａ₂）より大きく設定され、反力室４１の液圧は、反力ピストン２０の外径と入力軸１８のブレーキペダル側の外径との段部に、入力軸１８の入力と同方向に作用するようになっているが、図１６に示すように、この第９例のブレーキ液圧倍力装置１は、反力ピストン２０が省略されており、したがって入力軸１８のパワーピストン８側の外径（受圧面積Ａ₁）がブレーキペダル側の外径（受圧面積Ａ₂）より小さく設定され、反力室４１の液圧は、入力軸１８の段部１８ａに、入力軸１８の入力と逆方向に作用するようになっている。
【０１５８】
また、制御圧導入口４５に接続される可変サーボ装置７３は、ジャンピング作用を行うための第１切換弁１０７と、この第１切換弁１０７に直列に配設されて、逆二段サーボ作用を行うための第２切換弁１０８とから構成されている。
【０１５９】
第１切換弁１０７は二位置三方弁からなり、動力室２５の液圧によるパイロット圧で制御されるようにされている。そして、この第１切換弁１０７は、第２切換弁１０８をリザーバ３３に接続する第１位置Iと、第２切換弁１０８を出力口２７に接続する第２位置IIとが設定されており、ブレーキ液圧倍力装置１の非作動時は第１位置Iに設定されているとともに、動力室２５の液圧が、ジャンピング作用を終了させる第１設定作動圧となると、第２位置IIに切り換え設定されるようになっている。
【０１６０】
また、第２切換弁１０８も同様に二位置三方弁からなり、動力室２５の液圧によるパイロット圧で制御されるようにされている。そして、この第２切換弁１０８は、制御圧導入口４５を第１切換弁１０７に接続する第１位置Iと、制御圧導入口４５をリザーバ３３に接続する第２位置IIとが設定されており、通常時は第１位置Iに設定されているとともに、動力室２５の液圧が、サーボ比切換点の液圧に設定された第２設定作動圧になると第２位置IIに切り換え設定されるようになっている。
この第９例のブレーキ液圧倍力装置１の他の構成は、第５例と同じである。
【０１６１】
このように構成された第９例のブレーキ液圧倍力装置１においては、非作動時は、ブレーキ液圧倍力装置１は図示の状態になっている。この状態で、通常ブレーキ操作により、動力室２５に液圧が導入されると、パワーピストン８が前進する。このとき、動力室２５の液圧は、第１切換弁１０７が第１位置Iに設定されているため、反力室４１には導入されない。これにより、ブレーキ液圧倍力装置１は、前述の例と同様に通常ブレーキ時のサーボ比より大きなサーボ比でサーボ制御を行うようになる。このときの、ブレーキ液圧倍力装置１の入出力特性は、前述の数式１で与えられるとともに、図１２に示すように直線α″で表される。したがって、ブレーキ液圧倍力装置１はこの直線α″のサーボ比によるジャンピング作用を行うようになる。
【０１６２】
動力室２５の液圧が第１設定作動圧になると、第１切換弁１０７が作動して第２位置IIに切換設定され、反力室４１が出力口２７に接続され、動力室２５の液圧が反力室４１に導入される。これにより、サーボ比が小さくなって通常ブレーキ時のサーボ比と同じになり、これ以後、ブレーキ液圧倍力装置１は、この通常ブレーキ時のサーボ比でサーボ制御を行うようになる。このときの、ブレーキ液圧倍力装置１の入出力特性は、数式６で与えられるとともに、図１２に示すように直線β″で表される。
【０１６３】
動力室２５の液圧が第２設定作動圧になると、第２切換弁１０８が作動して第２位置IIに切換設定され、反力室４１がリザーバ３３に接続される。すると、反力室４１の圧液がリザーバ３３に排出され、反力室４１は大気圧となる。このため、サーボ比が通常ブレーキ時のサーボ比より大きいサーボ比でのサーボ制御を行うようになる。このときの、ブレーキ液圧倍力装置１の入出力特性は、前述の数式１で与えられるとともに、図１２に示すように直線α″と同じ直線である直線γ″で表される。
【０１６４】
このように、この第９例のブレーキ液圧倍力装置１においては、ブレーキ作動初期段階で、直線α″に沿った大きなサーボ比のサーボ制御によるジャンピング作用が行われ、その後直線β″に沿った小さなサーボ比のサーボ制御による通常ブレーキ作用が行われ、入力がある程度大きくなると、直線γ″に沿った大きなサーボ比のサーボ制御による、大きなブレーキ力を必要とするブレーキ制御が行われるようになる。こうして、第９例のブレーキ液圧倍力装置１は反力ピストン２０を有さなくても、第１切換弁１０７によってジャンピング特性を有するようになるとともに、第２切換弁１０８によって逆二段サーボ特性を有するようになる。
この第９例のブレーキ液圧倍力装置１の他の作用効果は、第５例と同じである。
【０１６５】
図１７は、本発明の実施の形態の第１０例を部分的に示す図である。
図１６に示す第９例では、反力室４１に対する圧液の給排を２つの切換弁１０７,１０８によって制御しているが、この第１０例のブレーキ制御装置１の可変サーボ装置７３では、図１７に示すようにこれらの切換弁１０７,１０８に代えて、電磁切換弁１０９が設けられている。また、可変サーボ装置７３は、ピストン９６、シリンダ９７、およびスプリング９８を有し、液を溜めるための低圧アキュムレータ９９と、指示圧導入口４５と電磁切換弁１０９および低圧アキュムレータ９９との間の通路に設けられたオリフィス１００と、このオリフィス１００をバイパスして設けられ、指示圧導入口４５から電磁切換弁１０９および低圧アキュムレータ９９への液の流れのみを許容するチェックバルブ１０１とを備えている。
【０１６６】
電磁切換弁１０９は、非作動時に制御圧導入口４５をリザーバ３３に接続する第１位置Iと、作動時に制御圧導入口４５を出力口２７に接続する第２位置IIとが設定されている。そして、電子制御装置は、圧力センサ６２によって検出された動力室２５の液圧が従来のジャンピング作用を終了するジャンピング終了圧になるまでは、この電磁切換弁１０９を非作動の第１位置Iに設定し、また動力室２５の液圧がジャンピング作用終了圧になったとき、電磁切換弁１０９を作動して第２位置IIに設定し、更に動力室２５の液圧がサーボ比切換点の圧力となったとき、電磁切換弁１０９を再び非作動の第１位置Iに設定するようになっている。
この第１０例のブレーキ液圧倍力装置１の他の構成は、第９例と同じである。
【０１６７】
この第１０例のブレーキ液圧倍力装置１においては、ブレーキ非作動時は、反力室４１がリザーバ３３に接続されて大気圧となっており、この状態で、通常ブレーキ操作が行われると、ブレーキ液圧倍力装置１は、図１２に示す直線α″の大きなサーボ比によるサーボ制御、すなわちジャンピング作用を行うようになる。
【０１６８】
動力室２５の液圧がジャンピング作用を終了する大きさになると、圧力センサ６２の検出信号に基づいて、電子制御装置が電磁切換弁１０９を第２位置IIに切り換え、反力室４１に動力室２５の圧液が導入される。このとき、通常ブレーキ時であるので、圧液は通常速度で流動しオリフィス１００によって絞られない。これにより、ブレーキ液圧倍力装置１は、ジャンピング作用が遅れることなく終了し、図１２に示す直線β″の小さいサーボ比による通常のサーボ制御を行うようになる。動力室２５の液圧がサーボ比切換点になると、電磁切換弁１０９が再び第１位置Iに切り換えられ、反力室４１が再びリザーバ３３に接続されるので、動力室２５の圧液がリザーバ３３に排出され、反力室４１の液圧が大気圧となる。これにより、ブレーキ液圧倍力装置１は、再び図１２に示す直線α″と同じ直線γ″の大きいサーボ比によるサーボ制御を行うようになる。
【０１６９】
ところで、急ブレーキ操作が行われると、動力室２５の液圧が急速に上昇する。すると、電磁切換弁１０９が第２位置IIに切り換えられたとき、動力室２５の圧液が反力室４１へ急速に流動するようになるが、この圧液はオリフィス１００によって絞られるとともに低圧アキュムレータ９９に貯えられ、その結果低圧アキュムレータ９９が蓄圧されるようになる。このため、反力室４１の液圧上昇が遅れるので、その分、図１２に示す直線α″の大きなサーボ比から直線β″の小さなサーボ比への移行が遅くなる。つまり、サーボ比切換点が変わり、ジャンピング作用の終了が遅くなる。したがって、その分ブレーキ液圧倍力装置１は大きな出力を発生するようになり、急ブレーキ時により大きなブレーキ力を得ることができる。
【０１７０】
このように、この第１０例のブレーキ液圧倍力装置１においても、ジャンピング作用と逆二段サーボ作用とを行うようになる。
この第１０例のブレーキ液圧倍力装置１の他の作用効果は、第９例と同じである。
【０１７１】
図１８は、本発明の実施の形態の第１１例を部分的に示す図である。
図１８に示すように、この第１１例のブレーキ制御装置１の可変サーボ装置７３では、図１４に示す第７例と同様に２個の第３および第４電磁開閉弁１１０,１１０が設けられている。
【０１７２】
第３電磁開閉弁１１０は、第７例の第４電磁開閉弁１０６と同じ常開の開閉弁であり、また第４電磁開閉弁１１１は第７例の第３電磁開閉弁１０５と同じ常閉の開閉弁である。
この第１１例のブレーキ液圧倍力装置１の他の構成は、第９例と同じである。
【０１７３】
この第１１例のブレーキ液圧倍力装置１においては、ブレーキ非作動時は、反力室４１がリザーバ３３に接続されて大気圧となっており、この状態で、通常ブレーキ操作が行われると、ブレーキ液圧倍力装置１は、図１２に示す直線α″の大きなサーボ比によるサーボ制御、すなわちジャンピング作用を行うようになる。
【０１７４】
動力室２５の液圧がジャンピング作用を終了する大きさになると、圧力センサ６２の検出信号に基づいて、電子制御装置が第３および第４電磁切換弁１１０,１１１をともに第２位置IIに切り換え、反力室４１に動力室２５の圧液が導入される。これにより、ブレーキ液圧倍力装置１は、図１２に示す直線β″の小さいサーボ比による通常のサーボ制御を行うようになる。動力室２５の液圧がサーボ比切換点になると、第３および第４電磁切換弁１１０,１１１がともに再び第１位置Iに切り換えられ、反力室４１が再びリザーバ３３に接続されるので、動力室２５の圧液がリザーバ３３に排出され、反力室４１の液圧が大気圧となる。これにより、ブレーキ液圧倍力装置１は、再び図１２に示す直線α″と同じ直線γ″の大きいサーボ比によるサーボ制御を行うようになる。
【０１７５】
このように、この第１１例のブレーキ液圧倍力装置１においても、ジャンピング作用と逆二段サーボ作用とを行うようになる。
この第１１例のブレーキ液圧倍力装置１の他の作用効果は、第９例と同じである。
【０１７６】
図１９は、本発明の実施の形態の第１２例を部分的に示す図である。
図１９に示すように、この第１２例のブレーキ制御装置１の可変サーボ装置７３では、図６に示す電磁比例制御弁６３および圧力調整弁６４が設けられており、アキュムレータ４０のアキュムレータ圧が圧力調整弁６４によって調整され、更に電磁比例制御弁６３によって制御されて反力室４１に導入されるようになっている。
この第１２例のブレーキ液圧倍力装置１の他の構成は、第９例と同じである。
【０１７７】
この第１２例のブレーキ液圧倍力装置１においては、非作動時は、電磁比例制御弁６３が反力室４１をリザーバ３３に接続し、反力室４１は大気圧となっている。この状態で、通常ブレーキ操作が行われると、ブレーキ液圧倍力装置１は、図１２に示す直線α″の大きなサーボ比によるサーボ制御、すなわちジャンピング作用を行うようになる。
【０１７８】
動力室２５の液圧がジャンピング作用を終了する大きさになると、圧力センサ６２の検出信号に基づいて、電子制御装置は電磁比例制御弁６３を作動制御するので、電磁比例制御弁６３は動力室２５の液圧に応じた圧力を出力し、この出力圧の圧液が反力室４１に導入される。これにより、ブレーキ液圧倍力装置１は、図１２に示す直線β″の小さいサーボ比による通常のサーボ制御を行うようになる。動力室２５の液圧がサーボ比切換点になると、電磁比例制御弁６３が非作動となり、反力室４１が再びリザーバ３３に接続されるので、動力室２５の圧液がリザーバ３３に排出され、反力室４１の液圧が大気圧となる。これにより、ブレーキ液圧倍力装置１は、再び図１２に示す直線α″と同じ直線γ″の大きいサーボ比によるサーボ制御を行うようになる。
【０１７９】
このように、この第１２例のブレーキ液圧倍力装置１においても、ジャンピング作用と逆二段サーボ作用とを行うようになる。
この第１２例のブレーキ液圧倍力装置１の他の作用効果は、第９例と同じである。
【０１８０】
なお、小さいサーボ比から大きいサーボ比へ移行する際に、電磁比例制御弁６３を適宜制御することにより、前述の図７および図１０に直線δ,δ′で示すと同様に滑らかに移行することができる。また、反力室４１に導入される液圧として、アキュムレータの蓄圧に代えて動力室２５の液圧を用いることもできる。
前述の実施例では、本発明の液圧倍力装置をブレーキ液圧倍力装置に適用して説明しているが、ブレーキ以外の他の液圧倍力装置にも適用できる。
【０１８１】
【発明の効果】
以上の説明から明らかなように、本発明の液圧倍力装置によれば、簡単な構造で、逆二段サーボ特性を得ることができるとともに、ジャンピング特性も得ることができるようになる。これにより、所定以上の入力で、通常の出力より大きな出力を得ることができる。
【０１８２】
また、サーボ比切換点を変更することができ、これにより種々の入出力特性に柔軟に対応することが可能となる。
更に、電磁比例制御弁を用いているので、サーボ比を滑らかに変えることができるようになる。
【０１８３】
更に、本発明のブレーキ制御システムによれば、簡単構造で、ジャンピング特性によりブレーキ力の立ち上がりを早くできるとともに、逆二段サーボ制御により、急ブレーキ時に迅速に大きなブレーキ力を得ることができるとともに、初心者等の運転者でも大きなブレーキ力を確実に得られるように補助することができる。
【図面の簡単な説明】
【図１】 本発明に係るブレーキ液圧倍力装置の実施の形態の第１例を示す断面図である。
【図２】 図１に示すブレーキ液圧倍力装置の部分拡大断面図である。
【図３】 図１に示すブレーキ液圧倍力装置に用いられている反力ピストンを示す断面図である。
【図４】 図１に示すブレーキ液圧倍力装置の入出力特性を示す図である。
【図５】 本発明の実施の形態の第２例を示す断面図である。
【図６】 本発明の実施の形態の第３例を示す断面図である。
【図７】 図６に示すブレーキ液圧倍力装置の入出力特性を示す図である。
【図８】 本発明の実施の形態の第４例を示す断面図である。
【図９】 図８に示すブレーキ液圧倍力装置の入出力特性を示す図である。
【図１０】図８に示すブレーキ液圧倍力装置の変形例の入出力特性を示す図である。
【図１１】本発明の実施の形態の第５例を示す断面図である。
【図１２】本発明の実施の形態の第５ないし第１１例の入出力特性を示す図である。
【図１３】本発明の実施の形態の第６例を示す断面図である。
【図１４】本発明の実施の形態の第７例を示す断面図である。
【図１５】本発明の実施の形態の第８例を示す断面図である。
【図１６】本発明の実施の形態の第９例を示す断面図である。
【図１７】本発明の実施の形態の第１０例を示す断面図である。
【図１８】本発明の実施の形態の第１１例を示す断面図である。
【図１９】本発明の実施の形態の第１２例を示す断面図である。
【図２０】従来のブレーキ液圧倍力装置を部分的に示す部分断面図である。
【図２１】図２０に示すブレーキ液圧倍力装置の入出力特性を示す図である。
【符号の説明】
１…ブレーキ液圧倍力装置、２…マスタシリンダ、３…ハウジング、８…パワーピストン、１８…入力軸、１８ａ…段部、２０…反力ピストン、２０ｅ…反力ピストンの右端、２１…スプリング、２５…動力室、２７…出力口、３３…リザーバ、４０…アキュムレータ、４１…反力室、４４…マスタシリンダピストン、４５…制御圧導入口、４６…切換弁、６０…制御弁、６１,９５,１０９…電磁切換弁、６２…圧力センサ、６３…電磁比例制御弁、６４…圧力調整弁、７３…可変サーボ装置、７４,１０７…第１切換弁、７５,１０８…第２切換弁、９９…低圧アキュムレータ、１００…オリフィス、１０１…チェックバルブ、１０５…第１電磁開閉弁、１０６…第２電磁開閉弁、１１０…第３電磁切換弁、１１１…第４電磁切換弁[0001]
BACKGROUND OF THE INVENTION
The present invention belongs to the technical field of a hydraulic booster that boosts an input to a predetermined magnitude by a hydraulic pressure and outputs the same, and a brake hydraulic booster that uses this hydraulic booster, and in particular, a servo. The present invention belongs to the technical field of a hydraulic booster that enables a servo ratio to be changed during control and a brake hydraulic booster that uses this hydraulic booster.
[0002]
[Prior art]
A hydraulic booster used in a brake hydraulic booster of an automobile or the like obtains a large output with a small input. As an example of this hydraulic booster, a brake hydraulic booster used in an automobile brake system has been proposed by a microfilm disclosed in Japanese Utility Model Application No. 4-33402 (Japanese Utility Model Application No. 5-84553).
[0003]
FIG. 20 is a diagram showing a brake hydraulic pressure booster disclosed in this microfilm. In the figure, 1 'is a brake hydraulic pressure booster, 2' is a housing, 3 'is a plug, 4' is a power piston, 5 'is a control valve, 6' is a valve seat member, 7 'is a cylindrical fixing member, 8 'is a nut, 9' is a ball valve, 10 'is a valve body, 11' is a cylindrical member, 12 'is an input shaft, 13' is a cylindrical stopper member, 14 'is a reaction force piston, and 15' is a power chamber. , 16 'is an output shaft.
[0004]
In the brake hydraulic pressure booster 1 ', in the illustrated non-operating state, the ball valve 9' of the control valve 5 'is seated on the valve seat member 6' and the tip valve portion of the tubular member 11 '. Is separated from the ball valve 9 '. Therefore, the power chamber 15 'is disconnected from the input port 17' that is always connected to a hydraulic pressure source (not shown), and communicates with the chamber 18 'that is always connected to a reservoir (not shown). No hydraulic pressure is introduced into ′, and the power piston 4 ′ does not operate.
[0005]
When input is applied from this non-operating state and the input shaft 12 'advances, the cylindrical member 11' also advances, and the tip valve portion of the cylindrical member 11 'contacts the ball valve 9' of the control valve 5 '. At the same time, the ball valve 9 'is pushed to be separated from the valve seat member 6'. As a result, the power chamber 15 'communicates with the input port 17' and is cut off from the chamber 18 ', pressure fluid is introduced into the power chamber 15', and the power piston 4 'operates. By the operation of the power piston 4 ′, the brake hydraulic pressure booster 1 ′ outputs from the output shaft 16 ′, operates the piston of a master cylinder (not shown), and the master cylinder generates brake hydraulic pressure. When the hydraulic pressure in the power chamber 15 ′ becomes a magnitude corresponding to the input, the ball valve 9 ′ is seated on the valve seat member 6 ′, so that the output of the brake hydraulic pressure booster 1 ′ is a magnitude obtained by boosting the input. It becomes.
[0006]
The reaction force piston 14 'is pressed backward against the spring 19' by the hydraulic pressure in the power chamber 15 ', but the hydraulic pressure in the power chamber 15' is still small, and the loss stroke of the brake system is not eliminated. In an initial stage where no braking force is generated, the reaction force piston 14 'does not come into contact with the step 12'a of the input shaft 12', so that a jumping action is performed by servo control with a very large boost ratio, that is, a servo ratio. . After the hydraulic pressure in the power chamber 15 ′ becomes a predetermined pressure and the reaction force piston 14 ′ comes into contact with the step portion 12 ′ a of the input shaft 12 ′, a braking force is substantially generated. Becomes a servo ratio of the normal brake, and thereafter, the brake hydraulic pressure booster 1 'performs servo control at the time of normal brake that generates an output obtained by boosting the input by this servo ratio.
[0007]
When the hydraulic pressure in the power chamber 15 ′ reaches the maximum pressure determined by the pressure generated by the hydraulic pressure source and does not increase any more, the brake hydraulic pressure booster 1 ′ becomes full load and does not perform servo control. After that, the output increase based on the input increase is not boosted.
[0008]
When the input is lost, the input shaft 12 'is retracted by a return spring (not shown), so that the cylindrical member 11' is also retracted, and the tip valve portion of the cylindrical member 11 'is separated from the ball valve 9' of the control valve 5 '. Sit down. As a result, the power chamber 15 'is shut off from the input port 17, communicated with the chamber 18', the hydraulic pressure introduced into the power chamber 15 'is discharged to the reservoir, and the power piston 4' is retracted by the return spring 20 '. To do. When the cylindrical stopper member 13 ′ fixed to the input shaft 12 ′ contacts the stopper 21 ′ of the plug 3 ′, the input shaft 12 ′ does not retreat further, but becomes a retreat limit, and the non-operating state shown in FIG. Return. When the hydraulic pressure in the power chamber 15 'is completely discharged, the power piston 4' also returns to the inactive state shown in the figure, the brake hydraulic pressure booster 1 'does not output, and the master cylinder is also inactive.
[0009]
In this conventional brake hydraulic pressure booster 1 ', the servo ratio in the servo control in which the brake force is substantially generated is constant as shown in FIG. Normally, this servo ratio is set according to the input / output characteristics desired during normal braking. If the servo ratio is constant in this way, the input / output characteristics are the same as in normal braking even during sudden braking, and the braking force is the same speed for the same input during both normal braking and sudden braking. It has come to rise.
[0010]
[Problems to be solved by the invention]
By the way, in a vehicle brake system, it is desirable to be able to generate a large braking force as soon as possible during sudden braking as compared to during normal braking.
[0011]
Also, it is necessary to generate a large braking force during sudden braking, but some drivers who are not used to driving automobiles, such as beginners, cannot depress the brake pedal greatly and cannot generate a large braking force. In such a case, it is desirable that a driver who is not used to driving can be assisted to reliably generate a large braking force.
[0012]
However, in the conventional brake hydraulic pressure booster 1 ′, since the servo ratio in the servo control in which the brake operation is substantially performed is constant, it is possible to generate a large braking force at the time of sudden braking earlier than during normal braking. In addition to being unable to assist, it is difficult to assist a person who is not accustomed to driving so as to reliably generate a large braking force, and it is difficult to reliably meet the above-mentioned demand.
[0013]
The present invention has been made in view of such circumstances, and an object thereof is to provide a hydraulic booster capable of obtaining a larger output than a normal output with a predetermined structure or more with a simple structure. It is to be.
[0014]
Another object of the present invention is to provide a brake that can obtain a large braking force as soon as possible during sudden braking, and that can assist a person who is not used to driving to generate a large braking force without fail. It is to provide a hydraulic booster.
[0015]
[Means for Solving the Problems]
In order to solve the above-mentioned problems, a hydraulic booster according to the invention of claim 1 includes a hydraulic pressure source for generating hydraulic pressure, a reservoir for storing hydraulic fluid, a power piston for generating output, and the power piston. A power chamber facing the pressure receiving surface of the power source, and shuts off the power chamber from the hydraulic pressure source when not in operation and communicates with the reservoir, and shuts off the power chamber from the reservoir when in operation and communicates with the hydraulic pressure source. A control valve that introduces the hydraulic fluid of the hydraulic pressure source into the power chamber according to the operation thereof, and controls the operation of the control valve, and the control valve side portion has a small diameter and is opposite to the control valve side. The stepped input shaft having a step portion with a large diameter, and the outer diameter is set to be larger than the diameter of the portion on the opposite side of the control shaft side of the input shaft and is formed in a cylindrical shape, and the input shaft The small diameter part is slidably fitted, and the front end is the power And a reaction force piston whose rear end can be brought into contact with the step portion of the input shaft, and a rear end of the reaction force piston is always urged away from the step portion of the input shaft, and the power chamber A spring for abutting the rear end of the reaction force piston against the step portion of the input shaft when the hydraulic pressure of the input pressure is equal to or higher than a first predetermined pressure, and an operating means for operating the input shaft when input is applied, A reaction force chamber is formed so that the step of the input shaft is located. Cage When the hydraulic pressure in the power chamber is greater than or equal to a second predetermined pressure that is greater than the first predetermined pressure, a hydraulic fluid having a servo ratio control pressure is introduced into the reaction force chamber, and the hydraulic pressure in the power chamber is increased to the second predetermined pressure. A servo ratio control pressure control valve is provided for controlling the pressure fluid in the reaction force chamber to be discharged to the reservoir when the pressure is smaller than the pressure.
[0016]
According to a second aspect of the present invention, there is provided a hydraulic pressure source for generating hydraulic pressure, a reservoir for storing hydraulic fluid, a power piston for generating output, a power chamber facing a pressure receiving surface of the power piston, The power chamber is shut off from the hydraulic pressure source and communicated with the reservoir. During operation, the power chamber is shut off from the reservoir and communicated with the hydraulic pressure source, and the hydraulic fluid from the hydraulic pressure source is adapted to the operation. A control valve to be introduced into the power chamber, and a stepped input shaft that controls the operation of the control valve and has a stepped portion having a large diameter on the control valve side and a small diameter on the side opposite to the control valve side. And an operating means for operating the input shaft when an input is applied, and a reaction force chamber is formed so that the step portion of the input shaft is located. Cage In addition, when the hydraulic pressure in the power chamber is greater than or equal to a predetermined pressure, a hydraulic fluid having a servo ratio control pressure is introduced into the reaction force chamber, and when the hydraulic pressure in the power chamber is lower than the predetermined pressure, Is provided with a servo ratio control pressure control valve for controlling so as to be discharged into the reservoir.
[0017]
The invention according to claim 3 is characterized in that the servo ratio control pressure control valve comprises one switching valve or two on-off valves.
Further, the invention according to claim 4 is that the switching valve or the on-off valve is controlled by the hydraulic pressure of the power chamber or by an electromagnetic force excited according to the hydraulic pressure of the power chamber. It is a feature.
[0018]
Furthermore, the invention of claim 5 provides a hydraulic pressure source for generating hydraulic pressure, a reservoir for storing hydraulic fluid, a power piston for generating output, a power chamber facing the pressure receiving surface of the power piston, and when not operating. The power chamber is shut off from the hydraulic pressure source and communicated with the reservoir. During operation, the power chamber is shut off from the reservoir and communicated with the hydraulic pressure source, and the hydraulic fluid from the hydraulic pressure source is adapted to the operation. A control valve to be introduced into the power chamber, and a stepped input shaft that controls the operation of the control valve and has a stepped portion having a large diameter on the control valve side and a small diameter on the side opposite to the control valve side. And an operating means for operating the input shaft when an input is applied, and a reaction force chamber is formed so that the step portion of the input shaft is located. And This reaction chamber Before When the hydraulic pressure in the power chamber is lower than the first predetermined pressure and higher than the second predetermined pressure higher than the first predetermined pressure, the servo ratio control pressure is introduced, and the hydraulic pressure in the power chamber is first. A servo ratio control pressure control valve is provided for controlling the pressure fluid in the reaction force chamber to be discharged to the reservoir when the pressure is equal to or higher than a predetermined pressure and lower than a second predetermined pressure.
[0019]
Further, in the invention of claim 6, an orifice is provided in a passage for introducing hydraulic pressure of the power chamber for obtaining an operating pressure signal for switching control of the servo ratio control pressure control valve by the hydraulic pressure of the power chamber. It is characterized by being.
[0020]
Further, the invention of claim 7 is directed to a hydraulic pressure source for generating hydraulic pressure, a reservoir for storing hydraulic fluid, a power piston for generating output, a power chamber facing a pressure receiving surface of the power piston, The power chamber is shut off from the hydraulic pressure source and communicated with the reservoir. During operation, the power chamber is shut off from the reservoir and communicated with the hydraulic pressure source, and the hydraulic fluid from the hydraulic pressure source is adapted to the operation. A control valve to be introduced into the power chamber, and a stepped input shaft that controls the operation of the control valve and has a stepped portion having a small diameter on the control valve side and a large diameter portion on the side opposite to the control valve side. And an operating means for operating the input shaft when an input is applied, and a reaction force chamber is formed so that the step portion of the input shaft is located. Cage , This reaction force chamber Before When the hydraulic pressure in the power chamber is equal to or higher than the first predetermined pressure and lower than the second predetermined pressure that is higher than the first predetermined pressure, the hydraulic pressure of the servo ratio control pressure is introduced, and the hydraulic pressure in the power chamber is changed to the first pressure. A servo ratio control pressure control valve is provided for controlling so that the pressure fluid in the reaction force chamber is discharged to the reservoir when the pressure is smaller than a predetermined pressure and when the pressure is equal to or higher than the second predetermined pressure.
[0021]
Further, the invention of claim 8 is characterized in that at least an orifice is provided in a passage between the servo ratio control pressure control valve and the reaction force chamber, and a low pressure accumulator located on the servo ratio control pressure control valve side from the orifice. It is characterized by being provided. Furthermore, the invention according to claim 9 is characterized in that the servo ratio control pressure control valve is composed of two switching valves, and these switching valves are controlled by hydraulic pressure of the power chamber.
[0022]
Further, in the invention of claim 10, the servo ratio control pressure control valve is composed of a switching valve or two on-off valves, and the switching valve or the on-off valve is excited according to the hydraulic pressure of the power chamber. It is characterized by being controlled by.
Furthermore, the invention of claim 11 is characterized in that the servo ratio control pressure is a hydraulic pressure of the power chamber.
[0023]
Furthermore, the invention of claim 12 is an electromagnetic proportional control valve in which the servo ratio control pressure control valve is controlled in accordance with the hydraulic pressure of the power chamber, and the servo ratio control pressure is the hydraulic pressure of the power chamber or The hydraulic pressure of the hydraulic pressure source is a hydraulic pressure controlled by the electromagnetic proportional control valve.
[0024]
Furthermore, the brake hydraulic pressure boosting system according to the thirteenth aspect of the invention is controlled by the hydraulic booster according to any one of the first to twelfth aspects of the invention and the output of the hydraulic booster to reduce the brake hydraulic pressure. And a brake cylinder that generates a braking force by introducing a brake fluid pressure of the master cylinder.
[0025]
Further, the invention of claim 14 is a two-system brake system, the operation of which is controlled by the hydraulic booster according to any one of claims 1 to 12 and the output of the hydraulic booster. The master cylinder that generates, the brake cylinder of one system that generates the brake force by introducing the hydraulic pressure of the power chamber of the hydraulic booster, and the brake hydraulic pressure of the master cylinder are introduced And a brake cylinder of the other system that generates the braking force by the semi-full power brake.
[0026]
[Action]
In the hydraulic booster according to the first aspect of the present invention having such a configuration, when the hydraulic pressure in the power chamber during operation is smaller than the first predetermined pressure, the reaction force piston does not contact the step portion of the input shaft. The hydraulic booster performs a jumping action. In addition, after the reaction force piston comes into contact with the stepped portion of the input shaft and the jumping action ends, the reaction force chamber is at atmospheric pressure while the hydraulic pressure in the power chamber is lower than the second predetermined pressure. Is a small servo ratio during normal braking, and servo control is performed at this servo ratio. Further, when the hydraulic pressure in the power chamber is equal to or higher than the second predetermined pressure, the servo ratio control pressure is introduced into the reaction force chamber, so that the servo ratio control pressure is formed by the reaction force piston and the input shaft. Acts on the stepped portion in the same direction as the input of the input shaft, resulting in a large servo ratio and an increase in the output of the hydraulic booster.
[0027]
In this way, a step is formed on the input shaft, the step is positioned in the reaction force chamber, and the reaction force piston is fitted to the small diameter portion of the input shaft. The force device exhibits a so-called reverse two-stage servo characteristic that changes from a small servo ratio to a large servo ratio when the input becomes larger than a predetermined value.
In the invention of claim 2, the jumping action is not performed, but the reverse two-stage servo action is performed as in the invention of claim 1.
[0028]
Furthermore, in the invention of claim 5, when the hydraulic pressure of the power chamber during operation is smaller than the first predetermined pressure, the hydraulic pressure of the servo ratio control pressure is introduced into the reaction force chamber, and this servo ratio control pressure is applied to the input shaft. This step acts in the same direction as the input of the input shaft, resulting in a large servo ratio and an increase in the output of the hydraulic booster. As a result, the hydraulic booster performs a jumping action. Further, while the hydraulic pressure in the power chamber is equal to or higher than the first predetermined pressure and lower than the second predetermined pressure, the pressure fluid in the reaction force chamber is discharged to the reservoir and the reaction force chamber is at atmospheric pressure. Is a small servo ratio during normal braking, and servo control is performed at this servo ratio. Further, when the hydraulic pressure in the power chamber becomes equal to or higher than the second predetermined pressure, the hydraulic fluid having the servo ratio control pressure is again introduced into the reaction force chamber, resulting in a large servo ratio, and the output of the hydraulic booster increases.
[0029]
Further, in the seventh aspect of the invention, when the hydraulic pressure in the power chamber during operation is smaller than the first predetermined pressure, the reaction force chamber is connected to the reservoir and the hydraulic pressure in the reaction force chamber is atmospheric pressure. The servo ratio becomes large, and the output of the hydraulic booster becomes large. As a result, the hydraulic booster performs a jumping action. Further, while the hydraulic pressure in the power chamber is equal to or higher than the first predetermined pressure and lower than the second predetermined pressure, the servo ratio control pressure is introduced into the reaction force chamber, and the servo ratio control pressure is supplied to the step portion of the input shaft. In addition, since it acts in the opposite direction to the input of the input shaft, the servo ratio is usually a small servo ratio during braking, and servo control is performed with this servo ratio. Further, when the hydraulic pressure in the power chamber becomes equal to or higher than the second predetermined pressure, the reaction force chamber is connected to the reservoir again, and the pressure fluid in the reaction force chamber is discharged to the reservoir so that the reaction force chamber becomes atmospheric pressure. The servo ratio becomes a large servo ratio, and the output of the hydraulic booster increases.
[0030]
In this way, in the inventions of claims 5 and 7, the hydraulic pressure booster is configured with a simple structure by simply forming the step portion on the input shaft and positioning the step portion in the reaction force chamber. Characteristics and reverse two-stage servo characteristics.
[0031]
In particular, in the invention of claim 6, the hydraulic pressure from the power chamber is throttled by the orifice, and the increase in the operating pressure of the servo ratio control pressure control valve is delayed. In the invention of claim 8, the servo pressure of the servo ratio control pressure is reduced. Is restricted by the orifice and the increase of the hydraulic pressure in the reaction force chamber is delayed, so that the shift from the large servo ratio to the small servo ratio is delayed. This increases the output of the hydraulic booster. Thus, the servo ratio switching point is changed.
In the invention of claim 12, since the servo ratio control pressure to the reaction force chamber is controlled by the electromagnetic proportional control valve, the servo ratio can be switched smoothly.
[0032]
Furthermore, in the brake control system according to the inventions of claims 13 and 14, the brake force rises quickly due to the jumping characteristic due to the simple structure, and the reverse two-stage servo control makes the brake control system large even for sudden braking or beginners. Brake force can be obtained.
[0033]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
FIG. 1 is a sectional view showing a first example of an embodiment of a brake hydraulic pressure booster according to the present invention, and FIG. 2 is a partially enlarged sectional view of FIG.
[0034]
As shown in FIGS. 1 and 2, the brake hydraulic pressure booster 1 according to the first example is provided with a master cylinder 2 integrally, and includes a housing 3 common to the master cylinder 2.
[0035]
In the housing 3, a stepped hole 4 that is relatively long in the axial direction is formed so as to open at the right end in FIG. 1, and the small diameter portion 4 a of the stepped hole 4 has a constant cross-sectional area and is multiplied by the brake hydraulic pressure. The force device 1 extends to the master cylinder 2. The right end opening of the axial hole 4 is liquid-tightly closed by a plug 6 having an O-ring 5. The plug 6 has a stepped cylindrical protrusion 6 a, and the small diameter protrusion 6 b of the stepped cylindrical protrusion 6 a is positioned within the small diameter portion 4 a of the stepped hole 4 of the housing 3. The large-diameter protruding portion 6c of the stepped cylindrical protruding portion 6a is press-fitted into the small-diameter portion 4a, and the plug 6 is brought into contact with the stepped portion of the stepped hole 4 by a nut 7 screwed into the housing 3. It is fixed to the housing 3.
[0036]
A power piston 8 is disposed in the small diameter portion 4a of the stepped hole 4 in a liquid-tight and slidable manner. The power piston 8 is provided with a stepped hole 9 that is positioned in the center and extends in the axial direction and opens at the right end of the power piston 8. The small diameter portion 9 a of the stepped hole 9 includes an end portion. A cylindrical valve seat member 10 having a first valve seat 10a is press-fitted into the tube. The flange portion 10b at the right end of the valve seat member 10 is in contact with the step portion of the stepped hole 9, and is supported in the axial direction by the cylindrical fixing member 11 fitted into the large diameter portion 9b of the stepped hole 9. Further, the cylindrical fixing member 11 is fixed to the power piston 8 by a C ring 12.
[0037]
A collar 13 is press-fitted into the small-diameter portion 9a of the stepped hole 9, and a cylindrical valve body 15 in which a conical valve 14 is integrally formed is slidably disposed in the collar 13. The valve body 15 is always urged by a spring 16 in the direction in which the conical valve 14 is seated on the first valve seat 10 a of the valve seat member 10. A second valve seat 17 a formed at the tip of the valve operating member 17 is disposed in the axial hole 10 c of the valve seat member 10 so as to be seated on the conical valve 14. Further, the valve operating member 17 is fitted and fixed to the input shaft 18, and the valve operating member 17 can be brought into contact with the tip of the small-diameter protruding portion 6 b of the plug 6. A flange-like stopper portion 17b that defines the retreat limit is integrally provided. A spring 19 is contracted between the valve seat member 10 and the valve actuating member 17, and the valve actuating member 17 and the input shaft 18 are always urged to the right in the drawing. The input shaft 18 penetrates the plug 6 in a liquid-tight manner, and its rear end is connected to a brake pedal (not shown).
[0038]
A cylindrical reaction force piston 20 is slidably fitted between the outer periphery of the input shaft 18 and the valve operating member 17 and the inner periphery of the axial hole of the small-diameter protruding portion 6b of the plug 6. Yes. As shown in FIG. 3, a first flange portion 20a and a second flange portion 20b are provided at the left end of the reaction force piston 20 in FIG. The left side portion of the first flange portion 20a can come into contact with the stopper portion 17b. The stopper portion 17b comes into contact with the left side portion of the first flange portion 20a, so that the valve against the reaction force piston 20 is provided. The stopper portion 20c prevents the operating member 17 from further retreating. In other words, when the stopper portion 17b of the valve operating member 17 comes into contact with the stopper portion 20c of the reaction force piston 20, further retreat of the input shaft 18 with respect to the reaction force piston 20 is prevented. Yes.
[0039]
Further, the right side portion of the second flange portion 20b is an engagement portion 20d that engages with the step portion 11a of the cylindrical fixing member 11 when the reaction force piston 20 moves backward by a predetermined amount with respect to the power piston 8. Yes. Furthermore, the right end 20 e of the reaction force piston 20 can come into contact with the step portion 18 a of the input shaft 18. A spring 21 is contracted between the second flange portion 20b of the reaction force piston 20 and the cylindrical fixing member 11. With this spring 21, the second flange portion 20b of the reaction force piston 20 is normally operated. The valve seat member 10 is in contact with the flange portion 10b.
[0040]
Further, the housing 3 is provided with an input port 22 through which pressure fluid is introduced, and a passage hole 23 that communicates the input port 22 with the small diameter portion 4a of the stepped hole 4. A passage hole 24 that communicates the passage hole 23 and the small diameter portion 9a of the stepped hole 9 is formed. In that case, the passage hole 24 opens to the small diameter portion 9 a between the valve seat member 10 and the collar 13. These input port 22 and passage holes 23 and 24 constitute a hydraulic pressure supply passage.
[0041]
A power chamber 25 is formed between the plug 6 and the right end of the power piston 8, and this power chamber 25 is always in communication with the axial hole 10 c of the valve seat member 10. In the power chamber 25, the stopper member 17b of the valve operating member 17 and the first and second flange portions 20a and 20b of the reaction force piston 20 are respectively positioned. A gap is provided between the outer peripheral surface of the small-diameter protruding portion 6 b of the plug 6 and the inner peripheral surface of the cylindrical fixing member 11, so that hydraulic fluid can freely flow on both sides in the axial direction of the cylindrical fixing member 11. It can flow.
[0042]
The power chamber 25 is always in communication with an output port 27 through a passage hole 26 formed in the housing 3, and the output port 27 is connected to the wheel cylinders 28, 1 in one of the two brake systems. 29 is always in communication.
[0043]
Further, an axial hole 30 penetrating in the axial direction is formed in the valve body 15, and this axial hole 30 is always in communication with a radial hole 31 formed in the power piston 8. The radial hole 31 is always in communication with a discharge port 32 formed in the housing 3 via the small diameter portion 4 a, and the discharge port 32 is always in communication with a reservoir 33.
[0044]
Further, the power chamber 25 is always in communication with a chamber 35 facing the step portion 15 a of the valve body 15 through a passage hole 34 formed in the power piston 8.
[0045]
Furthermore, a hydraulic pressure circuit 38 that connects the input port 22 and the reservoir 33 is provided with a hydraulic pressure pump 38 that is driven by a motor 37, and an accumulator 40 that is provided on the discharge side of the hydraulic pressure pump 38 via a check valve 39. It has been. A predetermined pressure is always stored in the accumulator 40 by the discharge pressure of the hydraulic pump 38.
[0046]
By the way, the brake hydraulic pressure booster 1 of this example is further provided with a reaction force chamber 41 formed in the plug 6, and in this reaction force chamber 41, the step portion 18 a of the input shaft 18 and the reaction force piston 20. The right end 20e of the head faces. The reaction force chamber 41 introduces control pressure through a radial hole 42 formed in the plug 6, an annular space 43 between the housing 3 and the plug 6, and an axial hole 44 formed in the housing 3. The port 45 is always in communication.
[0047]
As shown in FIG. 1, the control pressure introduction port 45 is connected to a pressure switching valve 46 composed of a two-position three-way valve. The two-position / three-way switching valve 46 includes a first position I that connects the control pressure introduction port 45 to a hydraulic circuit 36 that is always in communication with the reservoir 33, a control pressure introduction port 45, an output port 27, a wheel cylinder 28, 29 is set to the second position II connected to the brake fluid passage, and is normally set to the first position I, and the hydraulic pressure of the output port 27, that is, the hydraulic pressure of the power chamber 25 is set. When the set operating pressure is reached, the control is switched to the second position II.
[0048]
On the other hand, a master cylinder piston 47 of the master cylinder 2 is integrally formed at the front portion of the power piston 8, and this master cylinder piston 47 is set to have the same effective pressure receiving area as that of the power piston 8. The cup 3 is slidably fitted in the small diameter portion 4a of the stepped hole 4 of the housing 3 in a liquid-tight manner only in one direction.
[0049]
In addition, a liquid chamber 49 is defined in the small diameter portion 4a by the master cylinder piston 47, and this liquid chamber 49 is connected to the wheel cylinders 51, 52 in the other of the two brake systems via the output port 50. Is always in communication. Further, the master cylinder piston 47 is formed with an axial hole 53 that communicates the radial hole 31, that is, the reservoir 33 and the liquid chamber 49. A valve rod 55 provided with a valve 54 at its tip passes through the axial hole 53. The valve rod 55 is a valve release rod provided in the housing 3 through the small diameter portion 4a in the radial direction. 56 can be contacted. Further, the valve 54 is always urged by a spring 57 in a direction in which the valve 54 is seated on the valve seat 58. When the master cylinder piston 47 is in the inoperative position shown in the drawing, the valve rod 55 comes into contact with the valve release rod 56, so that the valve 54 is separated from the valve seat 58 against the spring force of the spring 57. The reservoir 33 and the liquid chamber 49 are communicated with each other. When the master cylinder piston 47 advances, the valve 54 is seated on the valve seat 58 and the valve rod 55 is separated from the valve release rod 56 by the spring force of the spring 57, and the reservoir 33 and the liquid chamber 49 are shut off. Master cylinder pressure is generated.
[0050]
Further, a return spring 59 that urges the power piston 8 integrated with the master cylinder piston 47 in the non-operation direction at all times is provided in the liquid chamber 49.
[0051]
Thus, in the brake system using the brake hydraulic pressure booster 1 integrated with the master cylinder 2 of this example, the hydraulic pressure of the power chamber 25 is introduced into the wheel cylinders 28 and 29 in one brake system. It is a power brake system, and the other brake system is configured as a semi-full power brake system that is a hydraulic brake system in which a master cylinder pressure is introduced into the wheel cylinders 51 and 52.
[0052]
Next, the operation of the brake hydraulic pressure booster integrated with the master cylinder of this example will be described.
When the brake is not operated when the brake pedal is not depressed, the conical valve 14, the first valve seat 10a of the valve seat member 10, and the second valve seat 17a of the valve operating member 17 are in the positional relationship shown in FIGS. . That is, the conical valve 14 is seated on the first valve seat 10 a of the valve seat member 10, and the second valve seat 17 a of the valve operating member 17 is separated from the conical valve 14. In this state, the passage hole 24 that is always in communication with the input port 22 and the axial hole 10c of the valve seat member 10 are blocked, and is always in communication with the axial hole 10c of the valve seat member 10 and the discharge port 32. The axial hole 30 of the valve body 15 is in communication. Therefore, when the brake is not operated, the power chamber 25 is disconnected from the pump 38 and the accumulator 40 and communicates with the reservoir 33, and no pressure fluid is supplied to the power chamber 25.
[0053]
The right end 20 e of the reaction force piston 20 is separated from the step portion 18 a of the input shaft 18. Further, the stopper portion 17b of the valve operating member 17 is in contact with the small-diameter protruding portion 6b of the plug 6, and is moved away from the stopper portion 20c of the first flange portion 20a of the reaction force piston 20 from the stopper portion 20c. Is in position. Further, the pressure switching valve 46 is in the illustrated first position I, and the reaction force chamber 32 communicates with the reservoir 33.
[0054]
When the brake is operated by depressing the brake pedal, the input shaft 18 moves forward, the second valve seat 17a of the valve operating member 17 is seated on the conical valve 14, and the conical valve 14 is seated on the first valve seat 10a of the valve seat member 10. In this state, the passage hole 24 and the axial hole 10c of the valve seat member 10 communicate with each other, and the axial hole 10c of the valve seat member 10 and the axial hole 30 of the valve body 15 are blocked. The Therefore, the power chamber 25 is shut off from the reservoir 33 and communicates with the pump 38 and the accumulator 40, and the pressure fluid of the accumulator 40 is supplied to the power chamber 25. In this case, the brake hydraulic pressure booster 1 that selectively switches the power chamber 25 to the hydraulic pressure source or the reservoir 33 of the pump 38 and the accumulator 40 by the conical valve 14, the first valve seat 10a, and the second valve seat 17a. The control valve 60 is configured.
[0055]
When the pressure fluid introduced into the power chamber 25 reaches a pressure that overcomes the spring force of the return spring 59, the fluid piston 8 moves forward by this fluid pressure, and the brake fluid pressure booster 1 generates an output, and the master cylinder. The piston 47 moves forward, the valve 54 is seated on the valve seat 58, and a master cylinder pressure is generated in the liquid chamber 49. Then, the hydraulic pressure in the power chamber 25 is introduced into both wheel cylinders 28 and 29 of one system, and the master cylinder pressure is introduced into both wheel cylinders 51 and 52 of the other system, and the brakes of both systems are activated. To do. At this time, since the effective pressure receiving area of the power piston 8 on which the hydraulic pressure in the power chamber 25 acts is equal to the effective pressure receiving area of the master cylinder piston 44 received by the master cylinder pressure in the liquid chamber 49, the liquid in the power chamber 25 The pressure and the master cylinder pressure are balanced and equal. Accordingly, the hydraulic fluid having the same hydraulic pressure is supplied to each of the wheel cylinders 28, 29; 51, 52.
[0056]
Further, the reaction force piston 20 is displaced to the right with respect to the power piston 8 and the input shaft 18 against the spring force of the spring 21 by the hydraulic pressure in the power chamber 25, but the hydraulic pressure in the power chamber 25 is In the initial stage where the pressure is relatively small, the rear end 20 e of the reaction force piston 20 does not reach the stepped portion 18 a of the input shaft 18.
[0057]
Further, the pressure fluid in the power chamber 25 is also introduced into the chamber 35 through the axial passage hole 34, and the fluid pressure in the chamber 35 acts on the step portion 15 a of the valve body 15, thereby causing the valve body 15. Is biased in a direction opposite to the hydraulic pressure in the power chamber 25.
[0058]
In the initial stage of operation in which there is a loss stroke of each wheel cylinder 28, 29; 51, 52 and these wheel cylinders do not substantially generate a braking force, the right end 20e of the reaction force piston 20 is brought into contact with the step portion 18a of the input shaft 18. Since no contact is made, the input shaft 18 is not subjected to any force from the reaction force piston 20. Therefore, the input shaft 18 is subjected to a force by the hydraulic pressure in the power chamber 25 received by a relatively small effective pressure receiving surface of the valve operating member 17 at the tip, and this force is applied to the driver as a reaction force. Reportedly.
[0059]
When the reaction force of the input shaft 18 becomes equal to the input of the input shaft 18, the conical valve 14 is seated on both the first valve seat 10 a of the valve seat member 10 and the second valve seat 17 a of the valve operating member 17. 25 is blocked from both the accumulator 40 and the reservoir 33. When the input of the input shaft 18 is further increased, the conical valve 14 is separated from the first valve seat 10a again, and the hydraulic fluid is further supplied to the power chamber 25, and the hydraulic pressure in the power chamber 25 is further increased. Thereafter, the conical valve 14 repeats seating and separation with respect to the first valve seat 10a, so that the hydraulic pressure in the power chamber 25 rises as the input of the input shaft 18 rises. Since the right end 20e of 20 is not in contact with the step portion 18a of the input shaft 18, the effective pressure receiving area of the input shaft 18 on which the hydraulic pressure in the power chamber 25 acts is small, and therefore the boost ratio, that is, the servo ratio at this time Is big. For this reason, the output of the brake hydraulic pressure booster 1 is greatly increased with respect to the input of the input shaft 18 with a large servo ratio, and the brake hydraulic pressure booster 1 performs a so-called jumping action.
[0060]
Now, the output of the brake hydraulic booster 1 is F _out The input of the brake hydraulic pressure booster 1 is F _in , A is the cross-sectional area of the outer diameter of the small diameter portion of the stepped input shaft 18 (effective pressure receiving area; hereinafter, other cross-sectional areas are the same). ₁ (Illustrated in FIG. 2; hereinafter, other cross-sectional areas are the same), and the cross-sectional area of the outer diameter of the large-diameter portion of the input shaft 18 is A ₂ , The cross-sectional area of the outer diameter of the reaction force piston 20 is A _Three The seat cross-sectional area of the conical valve 14 and the second valve seat 17a is A _Four The cross-sectional area of the outer diameter of the power piston 8 is A _Five (Where A _Four <A ₁ <A ₂ <A _Three <A _Five ), The spring force of each spring 19, 21, 59 is SPG ₁ , SPG ₂ , SPG _Three Then, the input / output characteristics of the brake hydraulic pressure booster 1 at this time are as follows:
[0061]
[Expression 1]

[0062]
And is represented by a straight line α as shown in FIG.
[0063]
When the hydraulic pressure in the power chamber 25 further increases and the power piston 8 further moves forward and the loss stroke of each wheel cylinder 28, 29; 51, 52 is eliminated, each wheel cylinder 28, 29; And the brake is substantially activated. In this state, the right end 20e of the reaction force piston 20 comes into contact with the step portion 18a of the input shaft 18 due to the increased hydraulic pressure in the power chamber 25, and the reaction force piston 20 is input by the urging force due to the hydraulic pressure in the power chamber 25. The shaft 18 acts to counteract the input of the input shaft 18. Accordingly, the reaction force applied to the input shaft 18 is increased and the jumping operation is finished. Thereafter, the output of the brake hydraulic pressure booster 1 rises smaller than the loss stroke with respect to the input of the input shaft 18. . That is, since the brake hydraulic pressure booster 1 increases the reaction force, the brake hydraulic pressure booster 1 performs servo control to boost and output the input of the input shaft 18 with a relatively small servo ratio, and the hydraulic pressure in the power chamber 25 The hydraulic pressure corresponds to the servo ratio. The servo ratio at this time is set to the servo ratio during normal braking. Further, during this servo control, the hydraulic pressure in the power chamber 25 does not rise to the operating pressure of the pressure switching valve 46 until the input becomes a predetermined magnitude, so that the pressure switching valve 46 is set to the first position I. The reaction force chamber 41 remains connected to the reservoir 33.
[0064]
Each wheel cylinder 28, 29; 51, 52 generates a braking force that is boosted with respect to the input of the input shaft 18, and the brake is operated by this braking force. At this time, as described above, the hydraulic pressure in the power chamber 25 and the master cylinder pressure are balanced and equal to each other, and the brake forces generated by the wheel cylinders 33, 34; 58, 59 are also equal to each other. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are as follows:
[0065]
[Expression 2]

[0066]
And is represented by a straight line β as shown in FIG.
[0067]
When the input reaches a predetermined amount and the hydraulic pressure in the power chamber 25 becomes the operating pressure of the pressure switching valve 46, the pressure switching valve 46 is switched and set to the second position II. Then, the reaction force chamber 41 is connected to a brake fluid passage between the output port 27 and the wheel cylinders 28 and 29, and the reaction force chamber 41 has a fluid pressure of the output port 27 that is a servo ratio control pressure, that is, a power chamber. A hydraulic pressure of 25 is introduced. The hydraulic pressure introduced into the reaction force chamber 41 is in the same direction as the input applied to the input shaft 18 at a part of the right end 20 e of the reaction force piston 20 that is in contact with the step portion 18 a of the input shaft 18. To act on. For this reason, the reaction force acting on the input shaft 18 is reduced, and servo control with the servo ratio of the normal brake represented by the straight line β is completed. Thereafter, the output of the brake hydraulic pressure booster 1 is the input shaft. For 18 inputs, it rises more than during servo control during normal braking. That is, since the brake hydraulic pressure booster 1 has a small reaction force, the brake hydraulic pressure booster 1 performs servo control that boosts and outputs the input of the input shaft 18 with a relatively large servo ratio. The servo ratio at this time is larger than the servo ratio during normal braking. Thereby, each wheel cylinder 28,29; 51,52 each generate | occur | produces a braking force larger than the braking force at the time of a normal brake with respect to the input of the input shaft 18. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are as follows:
[0068]
[Equation 3]

[0069]
And is represented by a straight line γ as shown in FIG. As described above, the brake booster 1 has a reverse two-stage servo characteristic that performs servo control with a servo ratio larger than the servo ratio during normal braking when the input becomes larger than a predetermined value.
[0070]
Further, when the input rises and the hydraulic pressure in the power chamber 25 reaches the maximum set pressure that is accumulated in the accumulator 40, the hydraulic pressure in the power chamber 25 does not increase any further, and the brake hydraulic pressure booster 1 has a large servo. Servo control by the ratio is finished and the load is fully loaded. Therefore, thereafter, the output increase of the brake hydraulic pressure booster 1 does not boost the input increase.
[0071]
When the brake pedal is released by releasing the brake pedal, both the input shaft 18 and the valve actuating member 17 are retracted to the right, and the second valve seat 17a of the control valve 60 is separated from the conical valve 14, and the inside of the power chamber 25 Is the axial hole 10c of the valve seat member 10, the gap between the conical valve 14 and the second valve seat 17a, the axial hole 30, the radial hole 31 of the valve body 15, and the small diameter of the stepped hole 4. It is discharged to the reservoir 33 through the part 4 a and the discharge port 32. At this time, the input shaft 18 is largely retracted until the stopper portion 17b of the valve operating member 17 contacts the stopper portion 20c of the reaction force piston 20, so that the second valve seat 17a is greatly opened from the conical valve 14, and the power chamber The pressure fluid in 25 is quickly discharged.
[0072]
As the hydraulic fluid in the power chamber 25 is discharged, the hydraulic fluid in both the wheel cylinders 28 and 29 of one system is quickly discharged to the reservoir 33 through the power chamber 25 and the hydraulic pressure in both the wheel cylinders 28 and 29 is reduced. descend. On the other hand, due to the spring force of the return spring 59, the master cylinder piston 44 and the power piston 8 are rapidly retracted, so that both the hydraulic pressure in the liquid chamber 49 and the hydraulic pressures in the other wheel cylinders 51, 52 are reduced. When the valve rod 55 comes into contact with the valve opening rod 56, the valve 54 is separated from the valve seat 58 and the liquid chamber 49 is connected to the reservoir 33 with respect to the subsequent retraction of the master cylinder piston 47. For this reason, the pressure fluid in both the wheel cylinders 51 and 52 is quickly discharged to the reservoir 33 through the fluid chamber 49, and the fluid pressure in both the wheel cylinders 51 and 52 is further reduced. As a result, the brakes of both systems are quickly released.
[0073]
When the hydraulic pressure in the power chamber 25 falls below the set operating pressure of the pressure switching valve 46, the pressure switching valve 46 is switched to the first position I, and the reaction force chamber 41 is connected to the reservoir 33. When the hydraulic pressure in the power chamber 25 further decreases to a predetermined pressure, the reaction force piston 20 moves forward relative to the power piston 8 and the input shaft 18 by the spring force of the spring 21 and the flange portion 10b of the valve seat member 10. And the right end 20 e of the reaction force piston 20 is separated from the step portion 18 a of the input shaft 18.
[0074]
When the input shaft 18 is further retracted until the brake release is almost completed, the stopper portion 17b of the valve operating member 17 comes into contact with the tip of the small-diameter protruding portion 6b of the plug 6 so that the input shaft 18 and the valve operating member 17 are retracted. It stops and both the input shaft 18 and the valve actuating member 17 become the retreat limit. However, even if the retraction of the input shaft 18 and the valve operating member 17 is stopped, the power piston 8, the reaction force piston 20, the conical valve 14 and the valve seat member 10 all continue to retreat. For this reason, the stopper portion 17b of the valve operating member 17 is separated from the stopper portion 20c of the reaction force piston 20, and the conical valve 14 approaches the second valve seat 17a of the valve operating member 17.
[0075]
When the right end of the power piston 8 comes into contact with the plug 6, the retraction of the power piston 8 stops, the master cylinder piston 47 and the power piston 8 are brought into the non-operating position, and the brake is released quickly and completely. When the brake is released, the output of the brake hydraulic pressure booster 1 is decreased along the full load, the straight line γ, the straight line β, and the straight line alpha, as opposed to when the input is increased, as the input decreases. Become.
[0076]
In the non-operating position of the power piston 8, the conical valve 14 comes very close to the second valve seat 17a of the valve actuating member 17, and the gap between the conical valve 14 and the second valve seat 17a becomes extremely small and is just before seating. . Therefore, as soon as the brake pedal is depressed and the input shaft 18 and the valve actuating member 17 move forward, the second valve seat 17a is seated on the conical valve 14 and the conical valve 14 is separated from the first valve seat 10a of the valve seat member 10 immediately. Sit down. That is, the loss stroke for performing the switching operation of the control valve 60 becomes extremely small, and the brake operates quickly.
[0077]
In this way, the brake is quickly activated when the brake is operated, and the brake is quickly released when the brake operation is released, so that the brake fluid pressure generating device 1 is extremely responsive.
According to the brake hydraulic pressure booster 1 of this example, during the servo control during normal braking, when the hydraulic pressure in the power chamber 25 is a predetermined pressure, in other words, when the input from the brake pedal is greater than a predetermined magnitude. In addition, the reverse two-stage servo characteristic that performs servo control with a servo ratio larger than the servo ratio during normal braking can be exhibited. In that case, a reaction force chamber 41 is provided at a position where the step portion 18 a of the input shaft 18 and the right end 20 e of the reaction force piston 20 are located, and the hydraulic pressure of the power chamber 25 is set in the reaction force chamber 41 by the pressure switching valve 46. This inverse two-stage servo characteristic can be obtained with a simple structure of introducing the above.
[0078]
Thus, at the time of sudden braking, it is possible to obtain a large braking force quickly by simply depressing the brake pedal to the servo ratio switching point without depressing considerably as in the conventional case. In addition, even a driver who is not used to driving can be assisted to reliably generate a large braking force during sudden braking.
[0079]
Further, the servo ratio switching point can be changed by making various adjustments by making the set operating pressure of the pressure switching valve 46 variable.
[0080]
Note that the accumulator pressure of the accumulator 40 may be introduced instead of the hydraulic pressure introduced into the reaction force chamber 41 in place of the hydraulic pressure in the power chamber 25. In this case, the accumulator pressure is regulated by the pressure regulating valve and introduced into the reaction force chamber 41. Since the accumulator pressure is higher than the hydraulic pressure in the power chamber 25 during normal braking, a large servo ratio, that is, a large output can be obtained for the same input. This introduction of accumulator pressure will be specifically described in another example described later.
[0081]
FIG. 5 is a view similar to FIG. 1, showing a second example of the embodiment of the present invention. The same components as those in the first example are denoted by the same reference numerals, and detailed description thereof is omitted. In the description of each example below, the same components as those in the previous example are similarly denoted by the same reference numerals, but the components not denoted by the same reference numerals are the same as the corresponding components in the other examples.
[0082]
In the first example described above, when changing the servo ratio, the hydraulic pressure in the power chamber 25 is introduced into the reaction force chamber 41 by the pressure switching valve 46 that is controlled by the hydraulic pressure in the power chamber 25. In the brake hydraulic pressure booster 1 of the second example, an electromagnetic switching valve 61 comprising the same two-position three-way valve is provided in place of the pressure switching valve 46, and the hydraulic pressure of the output port 27, that is, the power chamber 25 is provided. A pressure sensor 62 for detecting the hydraulic pressure is provided. When the detection signal of the pressure sensor 62 is input to an electronic control unit (not shown), and the electronic control unit determines that the hydraulic pressure in the power chamber 25 is equal to or higher than the predetermined pressure based on the detection signal, the electromagnetic switching valve 61 is switched to the first position. Switch to 2 position II.
[0083]
The brake hydraulic pressure booster 1 of the second example also has the reverse two-stage servo characteristic shown in FIG. 4 which is the same as that of the first example.
Other configurations and operational effects of the brake hydraulic pressure booster 1 of the second example are the same as those of the first example.
[0084]
FIG. 6 is a view similar to FIG. 1, showing a third example of the embodiment of the present invention.
[0085]
In the second example described above, the supply and discharge of the hydraulic pressure in the power chamber 25 with respect to the reaction force chamber 41 is controlled by the electromagnetic switching valve 61. However, the brake hydraulic pressure booster 1 of this third example is used. 6, an electromagnetic proportional control valve 63 including a linear solenoid valve is provided in place of the electromagnetic switching valve 61 as shown in FIG. 6, and a pressure adjustment valve 64 is provided, and the accumulator pressure of the accumulator 40 is adjusted to the pressure adjustment valve 64. The adjusted hydraulic pressure is controlled by the electromagnetic proportional control valve 63 and is introduced into the reaction force chamber 41. In this case, when the electronic control unit determines that the hydraulic pressure in the power chamber 25 is equal to or higher than a predetermined pressure based on the detection signal from the pressure sensor 62, the magnitude of the detection signal from the pressure sensor 62, that is, the hydraulic pressure in the power chamber 25. The electromagnetic proportional control valve 63 is actuated by outputting a control signal having a magnitude proportional to. The electromagnetic proportional control valve 61 normally connects the reaction force chamber 41 to the reservoir 33, but when activated, the reaction force chamber 41 is shut off from the reservoir 33, and the accumulator pressure is applied to the reaction force chamber 41. A controlled hydraulic pressure is introduced in proportion to the magnitude of the control signal from the electronic control device.
Other configurations of the brake hydraulic pressure booster 1 of the third example are the same as those of the first and second examples described above.
[0086]
In the brake hydraulic pressure booster 1 of the third example configured as described above, based on the detection signal from the pressure sensor 62, the electronic control unit sets the hydraulic pressure in the power chamber 25 to the electromagnetic proportional control valve 63. When it is determined that the pressure is lower than the operating pressure, the electromagnetic proportional control valve 63 is not operated, and the one electromagnetic proportional control valve 61 connects the reaction force chamber 41 to the reservoir 33. Therefore, at this time, it is exactly the same as the first and second examples, and the input / output characteristics of the brake hydraulic pressure booster 1 are servo characteristics along the straight line α, that is, jumping characteristics, and servo characteristics during normal braking along the straight line β. It becomes. When it is determined that the hydraulic pressure in the power chamber 25 has reached the set operating pressure of the electromagnetic proportional control valve 63, the electromagnetic proportional control valve 63 is activated, and the electromagnetic proportional control valve 63 is activated to cause the reaction force chamber 41 to enter. While shutting off from the reservoir 33, a hydraulic pressure proportional to the hydraulic pressure in the power chamber 25 is introduced into the reaction force chamber 41 by adjusting the hydraulic pressure of the accumulator pressure. As a result, the servo ratio is changed to the larger one as in the first and second examples. Since the hydraulic pressure introduced into the reaction chamber 41 increases in proportion to the hydraulic pressure in the power chamber 25, the servo ratio also gradually increases in proportion. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are as follows:
[0087]
[Expression 4]

[0088]
The servo characteristics along the straight line δ shown in FIG. Where P ₂ Is the hydraulic pressure in the reaction force chamber 41.
[0089]
Hydraulic pressure P of reaction chamber 41 ₂ Is the hydraulic pressure P in the power chamber 25 ₁ , The brake hydraulic pressure booster 1 performs servo control with a large servo ratio along the straight line γ, and the input / output of the brake hydraulic pressure booster 1 at that time The characteristic is given by Equation 2.
[0090]
According to the brake hydraulic pressure booster 1 of the third example, switching from the servo ratio of the straight line β to the servo ratio of the straight line γ can be smoothly performed along the straight line δ. Further, since the accumulated pressure of the accumulator 40 is introduced into the reaction force chamber 41, the output of the brake hydraulic pressure booster 1 is increased compared to the case where the hydraulic pressure of the power chamber 25 is introduced into the reaction force chamber 41. it can.
Other functions and effects of the brake hydraulic pressure booster 1 of the third example are the same as those of the first and second examples.
[0091]
By controlling the hydraulic pressure in the reaction force chamber 41, it is possible to smoothly change the servo ratio of the straight line β directly to the full load as shown by a two-dot chain line in FIG. Also in this third example, instead of the accumulator pressure, the hydraulic pressure in the power chamber 25 can be controlled and introduced into the reaction force chamber 41 by the electromagnetic proportional control valve 63.
[0092]
FIG. 8 is a view similar to FIG. 2, showing a fourth example of the embodiment of the present invention.
In each of the brake hydraulic pressure boosters 1 of the first to third examples described above, the spring 21 for urging the reaction force piston 20 is provided so as to have a jumping characteristic, but the brake hydraulic pressure of the fourth example is provided. In the booster 1, the spring 21 that urges the reaction force piston 20 is deleted to make the reaction force piston 20 free so that there is no jumping characteristic at the initial stage of brake operation.
Other configurations of the brake hydraulic pressure booster 1 of the fourth example are the same as those of any of the first to third examples.
[0093]
In the brake hydraulic pressure booster 1 of the fourth example configured as described above, when the input shaft 18 moves forward and hydraulic pressure is introduced into the power chamber during brake operation, the hydraulic pressure in the power chamber The reaction force piston 20 immediately moves backward relative to the input shaft 18, and the right end 20 e of the reaction force piston 20 immediately contacts the step portion 18 a of the input shaft 18. Therefore, the outer diameter of the input shaft 18 on the power piston 8 side is larger than the outer diameter of the input shaft 18 on the power piston 8 side due to the outer diameter of the reaction force piston 20. Further, the brake hydraulic pressure booster 1 of the third example immediately starts servo control based on the servo ratio during normal braking. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are as follows:
[0094]
[Equation 5]

[0095]
And is represented by a straight line β ′ as shown in FIG. That is, the jumping characteristics at the initial stage of braking as in the first and second examples are not performed.
[0096]
As in the case of the first example, when the hydraulic pressure in the power chamber 25 reaches the set operating pressure of the pressure switching valve 46, the hydraulic pressure in the power chamber 25 is introduced into the reaction force chamber 41. Therefore, the servo ratio of the brake hydraulic pressure booster 1 is changed to a servo ratio larger than the servo ratio during normal braking. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are as follows:
[0097]
[Formula 6]

[0098]
And is represented by a straight line γ ′ as shown in FIG.
Other functions and effects of the brake hydraulic pressure booster 1 of the fourth example are the same as those of any of the first to third examples.
[0099]
In the case of the fourth example, the reaction force piston 20 is not always necessary and can be omitted. In this case, the cross-sectional area of the portion of the input shaft 18 facing the power chamber 25 (that is, the portion including the flange portion 17b of the valve actuating member 17) is larger than the cross-sectional area of the brake pedal side portion of the input shaft 8, and A pressure receiving portion where the hydraulic pressure of the reaction force chamber 41 acts on the input shaft 18 in the same direction as the input to the portion of the shaft 18 facing the reaction force chamber 41, that is, the power piston 8 side of the stepped input shaft 18. It is necessary to make the outer diameter of the input shaft 18 larger than the outer diameter of the input shaft 18 on the brake pedal side.
[0100]
In the brake hydraulic pressure booster 1 of the fourth example, an electromagnetic switching valve 61 can be used instead of the pressure switching valve 46 as in the second example. Further, as in the third example, using the pressure sensor 62 and the electromagnetic proportional control valve 63, the hydraulic pressure obtained by adjusting the accumulator pressure is introduced into the reaction force chamber 41 in accordance with the hydraulic pressure in the power chamber 25. Thus, the servo ratio can be changed smoothly. In this case, the input / output characteristics of the brake hydraulic pressure booster 1 are as follows:
[0101]
[Expression 7]

[0102]
And is represented by straight lines β ′, δ ′, and γ ′ as shown in FIG. Other functions and effects of the brake hydraulic pressure booster 1 in this case are the same as in the third example.
[0103]
FIG. 11 is a view similar to FIG. 1, showing a fifth example of the embodiment of the present invention.
In the brake hydraulic pressure booster 1 of the fifth example, the spring 21 is omitted and the jumping characteristic by the reaction force piston 20 is not provided as in the fourth example.
[0104]
In the brake hydraulic pressure booster 1 of the fifth example, the master cylinder 2 is configured as a tandem master cylinder 2 having a primary piston 47 'and a secondary piston 47 ". The primary piston 47' It is integrally provided at the front end of the power piston 8. Further, the secondary piston 47 ″ is configured in the same manner as the master cylinder piston 47 in each of the above examples except for the rear end.
[0105]
A distance restricting rod 65 for restricting the distance between the pistons 47 ′ and 47 ″ protrudes and is fixed toward the master cylinder piston 47, and a retainer 66 is fitted to the distance restricting rod 65 so as to be slidable in the axial direction. Further, a spring 67 is contracted between the retainer 66 and the front end of the primary piston 47 ', and the retainer 66 is always urged away from the primary piston 47'. The retainer 66 abuts against the head 65a of the distance regulating rod 65 and is further restricted from moving away from the primary piston 47 '.
[0106]
Since the radial hole 31 is provided in the secondary piston 47 ″, the axial hole 30 of the valve body 15 of the control valve 60 is a passage hole 68 formed in the power piston 8 and the small diameter portion 4 a of the stepped hole 4. And is always connected to the reservoir 33 through a passage hole 69 formed in the housing 3.
[0107]
The rear end of the secondary piston 47 ″ is normally brought into contact with the retainer 66 by the spring force of the return spring 59. A cup seal 70 is provided at each of the front end of the primary piston 47 ′ and the rear end of the secondary piston 47 ″. 71, and a liquid chamber 72 is defined in the small diameter portion 4a between the cup seals 70, 71. Both cup seals 70, 71 allow the flow of liquid from the outside of the liquid chamber 72 passing through the cup seals 70, 71 to the inside of the liquid chamber 72, but from the liquid chamber 72 passing through the cup seals 70, 71 to the outside of the liquid chamber 72. The liquid flow to the is prevented.
[0108]
The effective pressure receiving area of the power piston 8, the effective pressure receiving area of the primary piston 47 ', and the effective pressure receiving areas of the front and rear end portions of the secondary piston 47 "are all set equal.
[0109]
A variable servo device 73 is connected to the control pressure introduction port 45. The variable servo device 73 is arranged in series with the first switching valve 74 for giving the brake hydraulic pressure booster 1 a jumping characteristic, and the first switching valve 74. And a second switching valve 75 for providing reverse two-stage servo characteristics.
[0110]
The first switching valve 74 is a two-position three-way valve, and is controlled by a pilot pressure based on the hydraulic pressure in the power chamber 25. In the first switching valve 74, a first position I for connecting the second switching valve 75 to the output port 27 and a second position II for connecting the second switching valve 75 to the reservoir 33 are set. When the brake hydraulic pressure booster 1 is not in operation, it is set to the first position I, and the hydraulic pressure in the power chamber 25 urges the first set operating pressure (reaction force piston 20 to end the jumping action). In the above-described example having the spring 21, which corresponds to the operating pressure of the reaction force piston 20), the switching is set to the second position II.
[0111]
Similarly, the second switching valve 75 is a two-position three-way valve, and is controlled by a pilot pressure generated by the hydraulic pressure in the power chamber 25. The second switching valve 75 is
A first position I for connecting the control pressure introduction port 45 to the first switching valve 74 and a second position II for connecting the control pressure introduction port 45 to the output port 27 are set. And the hydraulic pressure in the power chamber 25 is switched to the second position II when the hydraulic pressure in the power chamber 25 reaches the second set operating pressure set at the hydraulic pressure at the servo ratio switching point.
[0112]
By the way, the brake hydraulic pressure booster 1 of the fifth example includes several devices in addition to the variable servo device 73 that changes the servo ratio.
That is, the liquid chamber 72 is connected to the variable stroke device 77 via the connection port 76 of the housing 3. The variable stroke device 77 includes a piston 78, a cylinder 79, and a spring 80, and includes a stroke simulator 81 for securing a pedal stroke, and an orifice 82 provided in a passage between the connection port 76 and the stroke simulator 81. And a check valve 83 that bypasses the orifice 82 and permits only the flow of liquid from the stroke simulator 81 to the connection port 76.
[0113]
Further, a connection port 84 is formed in the housing 3, and this connection port 84 is connected to a liquid passage 85 branched from the hydraulic circuit 36. An automatic brake device 86 is provided in the liquid passage 85, and the automatic brake device 86 includes an electromagnetic switching valve 87 formed of a two-position three-way valve and a pressure adjustment valve 88. In the electromagnetic switching valve 87, a first position I for connecting the connection port 84 to the reservoir 33 and a second position II for connecting the connection port 79 to the accumulator 40 via the pressure regulating valve 88 are set. During normal operation, the first position I is set, and when the automatic brake is activated, the second position II is set. In the inoperative position of the primary piston 47 ′, the cup seal 70 is positioned between the passage hole 69 and the connection port 84. Therefore, when not operating, the liquid flows freely between the liquid chamber 72 and the connection port 84 in both directions, but when the power piston 8 moves forward and the cup seal 70 passes the connection port 84, The flow of liquid from the connection port 84 toward the liquid chamber 72 is allowed, but the flow of liquid from the liquid chamber 72 toward the connection port 84 is blocked.
[0114]
In the fluid passage that connects the output port 27 of the brake hydraulic pressure booster 1 and the wheel cylinders 28 and 29 of one system, the brake of one system is braked when the hydraulic pressure source of the pump 38 and the accumulator 40 fails. A brake operating device 89 is provided at the time of hydraulic pressure failure for reliable operation. This hydraulic pressure failure brake operating device 89 includes a pressure conversion cylinder 93 having a piston 90, a cylinder 91, and a spring 92, and a pressure control switching valve 94 including a two-position three-way valve.
[0115]
When pressure fluid discharged from the pump 38 is introduced into the pressure conversion cylinder 93, the piston 90 operates to generate brake fluid pressure, and this brake fluid pressure is introduced into the wheel cylinders 28 and 29. ing. Further, in the pressure conversion cylinder 93, when the hydraulic pressure failure is caused by the failure on the side of the wheel cylinders 28 and 29, the liquid discharged from the pump 38 (that is, the accumulator pressure-accumulating liquid) goes out from the failed portion. It is designed to prevent leakage.
[0116]
The pressure control switching valve 94 is controlled by a pilot pressure generated by accumulating the accumulator 40. The pressure control switching valve 94 is set with a first position I for connecting the pressure conversion cylinder 93 to the output port 27 and a second position II for connecting the pressure conversion cylinder 93 to the connection port 76. The first position I is set when the pressure is normal, and the second position II is switched when the hydraulic pressure is lost. Other configurations of the brake hydraulic pressure booster 1 of the fifth example are the same as those of the fourth example.
[0117]
In the brake hydraulic pressure booster 1 of the fifth example configured as described above, the brake hydraulic pressure booster 1 is in the state shown in the drawing when the hydraulic pressure is normal and when it is not operating. In this state, when a normal brake operation is performed and hydraulic pressure is introduced into the power chamber 25, the power piston 8 moves forward, the reaction force piston 20 immediately operates, and its right end 20e is the step of the input shaft 18. Simultaneously with the contact with the portion 18a, the hydraulic pressure in the power chamber 25 is immediately introduced into the reaction force chamber 41. As a result, the brake fluid pressure booster 1 performs servo control with a servo ratio larger than the servo ratio during normal braking, as in the above example. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are given by the above-described equation 6 and are represented by a straight line α ″ as shown in FIG. 12. Therefore, depending on the servo ratio of the straight line α ″. The servo control performs the same jumping action as the jumping action by the reaction force piston 20.
[0118]
When the power piston 8 advances, the primary piston 47 ′ also advances, and when the cup seal 70 passes through the connection port 84, the liquid chamber 72 is blocked from the reservoir 33 and is in a sealed state. For this reason, as the primary piston 47 ′ moves forward, the liquid in the liquid chamber 72 is sent from the connection port 76 to the stroke simulator 81 of the variable stroke device 77. At this time, the primary piston 47 'moves forward at a normal speed because it is during a normal brake operation, so the orifice effect of the orifice 82 on the liquid flowing to the stroke simulator is small. Accordingly, the primary piston 47 ′, that is, the power piston 8, strokes at a normal speed by the stroke of the piston 78, that is, the stroke simulator of the stroke simulator 81.
[0119]
When the hydraulic pressure in the power chamber 25 reaches the first set operating pressure, the first switching valve 74 is activated and switched to the second position II, the reaction force chamber 41 is shut off from the output port 27 and connected to the reservoir 33. Is done. For this reason, the hydraulic pressure in the reaction force chamber 41 is discharged to the reservoir 33, and the servo ratio is reduced to be the same as the servo ratio during normal braking as in the above example. Thereafter, the brake hydraulic pressure booster 1 Will perform servo control with the servo ratio during normal braking. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are given by the above-described equation 5 and are represented by a straight line β ″ as shown in FIG.
[0120]
When the hydraulic pressure in the power chamber 25 reaches the second set operating pressure, the second switching valve 75 is actuated to switch to the second position II, the reaction force chamber 41 is shut off from the reservoir 33, and again to the output port 27. Connected. For this reason, the hydraulic pressure of the power chamber 25 is reintroduced into the reaction force chamber 41, and servo control is performed with a servo ratio larger than the servo ratio during normal braking, as in the above example. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are given by the above-described equation 6 and also represented by a straight line γ ″ which is the same straight line as the straight line α ″ as shown in FIG.
[0121]
Thus, in the brake hydraulic pressure booster 1 of the fifth example, a jumping action is performed by servo control with a large servo ratio along the straight line α ″ at the initial stage of brake operation, and then along the straight line β ″. When a normal braking action is performed by servo control with a small servo ratio and the input becomes large to some extent, brake control that requires a large braking force such as sudden braking is again performed by servo control with a large servo ratio along the straight line γ ″. To be done.
[0122]
Thus, even if the brake hydraulic pressure booster 1 of the fifth example does not have the jumping characteristic by the reaction force piston 20, it has a simple structure and has the jumping characteristic by the first switching valve 74, and the second The switching valve 75 has a reverse two-stage servo characteristic.
[0123]
By the way, the brake hydraulic pressure booster 1 of the fifth example not only performs the variable servo action by the jumping action and the reverse two-stage servo characteristic, but also the variable stroke action, the automatic brake action, and the brake when the hydraulic pressure is lost. The action is also performed.
[0124]
First, the operation of the variable stroke will be described. The case where the driver depresses the brake pedal at the normal braking speed is as described above. Further, when the brake pedal is rapidly depressed and the brake is suddenly applied, the power piston 8 and the primary piston 47 ′ also advance rapidly, so that the liquid in the liquid chamber 72 is rapidly fed from the connection port 76 to the stroke simulator 81. . At this time, since the flow velocity of the liquid from the connection port 76 is fast, the orifice effect by the orifice 82 is increased, and a high liquid pressure is generated in the liquid chamber 72, and the high liquid pressure in the liquid chamber 72 is applied to the primary piston 47 ′ and the power. Since it acts as a large reaction force on the input shaft 18 via the piston 8, the pedal stroke becomes smaller during normal braking. Therefore, the hydraulic booster 1 generates a large output due to the large reaction force, and the secondary piston 47 ″ of the master cylinder 2 generates a high master cylinder pressure due to the large output, and the secondary piston 47 ″ is effective. Since the pressure receiving area and the effective pressure receiving area of the power piston 8 are the same, the hydraulic pressure in the power chamber 25 becomes a high hydraulic pressure equal to the master cylinder pressure.
[0125]
Then, a high hydraulic pressure in the power chamber 25 is supplied to the pressure conversion cylinder 93 from the output port 27, and the piston 90 of the pressure conversion cylinder 93 is operated to generate a high brake hydraulic pressure. This high brake hydraulic pressure is applied to the wheel cylinder. 28 and 29, the wheel cylinders 28 and 29 generate a large braking force. On the other hand, a high master cylinder pressure is introduced from the output port 50 to the wheel cylinders 51 and 52, and the wheel cylinders 51 and 52 generate a large braking force.
[0126]
When the brake operation is released by releasing the brake pedal, the liquid supplied to the stroke simulator 81 is returned to the liquid chamber 72 without delay by the check valve 83. Therefore, even if the orifice 82 is provided, the power piston 8 and the input shaft 18 return to the non-operating position without delay.
[0127]
In this way, the pedal stroke can be changed by the orifice 82 according to the depression speed of the brake pedal. When the brake pedal is rapidly depressed, the hydraulic booster 1 is small in the input shaft 18 by the variable stroke device 77. Since a large output is generated by the stroke, the brake force rises quickly and a large brake force is quickly generated in both systems.
[0128]
Next, the automatic braking action will be described. When the automatic brake operation condition is satisfied while the vehicle is running, an electronic control device (not shown) switches the electromagnetic switching valve 87 to the second position II. For this reason, the connection port 84 is connected to the accumulator 40 via the pressure regulating valve 88. Then, the accumulated pressure of the accumulator 40 is adjusted to a predetermined pressure by the pressure adjusting valve 88, and this adjusted fluid pressure is introduced into the fluid chamber 72 through the connection port 84, and the pressure control switching valve 94 is caused by this fluid pressure. Switching to the second position II is set. The fluid pressure introduced into the fluid chamber 72 is further introduced into the variable stroke device 77 through the connection port 76 and also into the pressure conversion cylinder 93 of the brake operating device 89 when the fluid pressure is lost. Then, the piston 90 is activated to generate brake fluid pressure, and the generated brake fluid pressure is introduced into the wheel cylinders 28 and 29, so that the brake of one system is activated.
[0129]
On the other hand, the hydraulic pressure introduced into and adjusted in the liquid chamber 72 acts on the rear end face of the secondary piston 47 ″ of the master cylinder 2, so that the secondary piston 47 ″ operates and the liquid pressure 49 A master cylinder pressure equal to 72 hydraulic pressure is generated. This master cylinder pressure is introduced into the wheel cylinders 51 and 52, and the brake of the other system is operated. In this way, the automatic brake operates reliably in both systems.
[0130]
When the condition for releasing the operation of the automatic brake is satisfied, the electronic control unit sets the electromagnetic switching valve 87 to the inoperative first position I again. For this reason, since the connection port 76 is connected to the reservoir 33, the hydraulic pressure in the liquid chamber 72 and the pressure conversion cylinder 93 is discharged to the reservoir 33. At this time, when the hydraulic pressure at the connection port 84 decreases to some extent, the pressure control switching valve 94 is returned to the first position I, and the pressure conversion cylinder 93 is connected to the power chamber 25. Is discharged to the reservoir 33 through the power chamber 25. As a result, the brake fluid pressure in the pressure conversion cylinder 93 disappears, and the brake of one system is released.
[0131]
Further, when the fluid chamber 72 is connected to the reservoir 33 and the fluid pressure is reduced, the secondary piston 47 ″ is also retracted and the secondary piston 47 ″ returns to the non-operating position. Similarly, since the liquid chamber 49 communicates with the reservoir 33, the master cylinder pressure disappears and the brake of the other system is released. Thus, the automatic brake is completely released.
[0132]
Next, the braking action at the time of hydraulic pressure failure will be described. When the hydraulic pressure of the hydraulic pressure source such as the pump 38 and the accumulator 40 is lost, the pressure control switching valve 94 is set to the second position II. In this state, even if the driver performs a normal brake operation by depressing the brake pedal to advance the input shaft 18 and switch the control valve 60, no hydraulic pressure is introduced into the power chamber 25. For this reason, the power piston 8 does not operate depending on the hydraulic pressure in the power chamber 25. Further, when the brake pedal is greatly depressed and the input shaft 18 moves forward greatly, the valve body 15 makes a maximum stroke and comes into contact with the power piston 8 to push the power piston 8. Then, when the primary piston 47 ′ integrated with the power piston 8 moves forward and the cup seal 70 passes through the connection port 84, a hydraulic pressure is generated in the liquid chamber 72, and this hydraulic pressure is converted into a pressure via the connection port 76. It is introduced into the cylinder 93. Subsequent brakes in one system operate in the same manner as in the automatic brake described above.
[0133]
Further, in the brake operation when the hydraulic pressure is lost, the primary piston 47 'moves forward, the secondary piston 47 "moves forward, and the valve 54 is seated on the valve seat 58 in the same manner as described above. The hydraulic pressure in the fluid chamber 49 is introduced into the other wheel cylinders 51 and 52 via the output port 50, and the other brake system is also operated. Since the effective pressure receiving areas at the front and rear ends are equal, the fluid pressure in the fluid chamber 72 and the fluid pressure in the fluid chamber 49 are the same, and as a result, the braking force of both systems is the same.
[0134]
The release of the brake operation when the hydraulic pressure is lost is performed by releasing the brake pedal in the same manner as the release of the normal brake. The release of the brake pedal causes the primary piston 47 'to move backward together with the power piston 8 and the fluid pressure in the fluid chamber 72 decreases, so that the braking force of one system decreases and the secondary piston 47 "moves backward. Since the hydraulic pressure in the liquid chamber 49 decreases, the braking force of the other system also decreases, and when the primary piston 47 ′ moves backward and the cup seal 70 passes through the connection port 84, the liquid chamber 72 enters the connection port 84. Then, since the fluid chamber 72 communicates with the reservoir 33, the fluid pressure in the fluid chamber 72 and the pressure conversion cylinder 93 is discharged to the reservoir 33, so that the brake of one system is completely released. Since the master cylinder piston 47 is further retracted, the valve 54 is separated from the valve seat 58 in the same manner as the normal brake operation is released. Communicates with the server 33, the brake of the other system is also completely released.
Other effects of the brake hydraulic pressure booster 1 of the fifth example are the same as those of the second example.
[0135]
FIG. 13 is a view similar to FIG. 11, showing a sixth example of the embodiment of the present invention. The brake hydraulic pressure booster 1 of the sixth example is connected to the first and second switching valves 74 and 75 controlled by the hydraulic pressure of the power chamber 25 in the variable servo device 73 of the fifth example shown in FIG. Instead, an electromagnetic switching valve 95 comprising a two-position three-way valve is provided. The electromagnetic switching valve 95 has a first position I for connecting the control pressure introduction port 45 to the output port 27 when not operated and a second position II for connecting the control pressure introduction port 45 to the reservoir 33 during operation. Yes. Further, the hydraulic pressure introduction passage to the pressure sensor 62 includes an orifice 100 and a check valve 101 that is provided to bypass the orifice 100 and allows only the liquid flow in the direction discharged from the pressure sensor 62. Yes.
[0136]
Then, the electronic control unit turns off the electromagnetic switching valve 95 until the hydraulic pressure at the output port detected by the pressure sensor 62, that is, the hydraulic pressure in the power chamber 25 reaches the jumping end pressure for ending the conventional jumping action. When the first position I of the operation is set and the hydraulic pressure in the power chamber 25 reaches the jumping action end pressure, the electromagnetic switching valve 95 is operated to be set to the second position II, and further the hydraulic pressure in the power chamber 25 is set. When the pressure becomes the servo ratio switching point, the electromagnetic switching valve 95 is set to the non-actuated first position I again.
[0137]
The variable stroke device 72 is connected to the connection port 76 via the first electromagnetic on-off valve 102. The first electromagnetic open / close valve 102 has a communication position I and a cutoff position II, and is normally a normally open valve set to the communication position I. Further, the variable stroke device 72 includes a second electromagnetic on-off valve 103 provided in a passage connecting the connection port 76 and the pressure conversion cylinder 93. The second electromagnetic on-off valve 103 is set to a cutoff position I and a communication position II, and is normally a normally closed valve set to the cutoff position I.
[0138]
The pressure control switching valve 94 in the brake operating device 89 at the time of hydraulic pressure failure of the fifth example described above is controlled to be switched by the accumulated pressure of the accumulator 40. The brake at the time of hydraulic pressure failure of the sixth example is described. In the operating device 89, the pressure control switching valve 94 is an electromagnetic switching valve composed of a two-position three-way valve. A pressure sensor 104 that detects the accumulated pressure in the accumulator 40 is provided for switching control of the electromagnetic switching valve.
[0139]
When the electronic control unit determines that the automatic brake operation condition is satisfied, and determines that the accumulated pressure of the accumulator 40 has failed based on the detection signal from the pressure sensor 104, the electronic control unit sets the pressure control switching valve 94. Switching is set.
The other configuration of the brake hydraulic pressure booster of the sixth example is the same as that of the fifth example.
[0140]
The operation of the brake hydraulic pressure booster of the sixth example configured as described above will be described.
When the brake is not operated, the components of the brake hydraulic pressure booster 1 and the master cylinder 2 are in the non-operating position shown in FIG. In this state, when the normal brake operation is performed by depressing the normal speed of the brake pedal, the pressure fluid is introduced into the power chamber 25 as described above, and the pressure fluid in the power chamber 25 is further supplied to the output port 27, The reaction pressure chamber 41 is introduced from the control pressure introduction port 45 through the electromagnetic switching valve 95. At this time, since the brake pedal is depressed at a normal speed, the liquid is introduced into the reaction force chamber 41 with almost no restriction by the orifice 100. Thus, like the fifth example, the brake hydraulic pressure booster 1 performs servo control based on the servo ratio of the straight line α ″ shown in FIG. 12, that is, jumping action.
[0141]
When the hydraulic pressure in the power chamber 25 is large enough to complete the jumping action, the electronic control unit switches the electromagnetic switching valve 95 to the second position II based on the detection signal from the pressure sensor 62. Then, the reaction force chamber 41 is connected to the reservoir 33, and the pressure liquid introduced into the reaction force chamber 41 passes through the check valve 101 and the electromagnetic switching valve 95 and is discharged without delay to the reservoir 33. As a result, the brake hydraulic pressure booster 1 performs normal servo control with a small servo ratio of the straight line β ″ shown in FIG. 12. When the hydraulic pressure in the power chamber 25 becomes the servo ratio switching point, the electromagnetic switching is performed. The valve 95 is again switched to the first position I. Then, the reaction force chamber 41 is again connected to the output port 27, so that the pressure fluid in the power chamber 25 is introduced into the reaction force chamber 41 and the reaction force chamber 41 liquid. The pressure becomes the same as the hydraulic pressure in the power chamber 25. Thus, the brake hydraulic pressure booster 1 performs servo control with a servo ratio having a large straight line γ ″ shown in FIG.
[0142]
By the way, when the brake pedal is depressed much more rapidly than the normal speed and a sudden braking operation is performed, the hydraulic pressure in the power chamber 25 increases rapidly. As a result, the pressure fluid in the power chamber 25 rapidly flows into the reaction force chamber 41, but the increase in the fluid pressure of the pressure sensor 62 is delayed by the orifice 100, so that the switching of the electromagnetic switching valve 95 is delayed. Due to this delay, the transition from the large servo ratio of the straight line α ″ shown in FIG. 12 to the small servo ratio of the straight line β ″ is delayed. That is, the servo ratio switching point changes and the end of the jumping action is delayed. Therefore, the brake hydraulic pressure booster 1 accordingly generates a large output, and can obtain a larger braking force during sudden braking.
[0143]
On the other hand, in the variable stroke device 77, when each solenoid valve 94, 102, 103 is in the non-actuated position shown in FIG. 13, it operates in exactly the same way as in the fifth example described above, and the brake pedal operates as a stroke simulator 81. Stroke by the stroke simulator. Further, when only the first electromagnetic opening / closing valve 102 is operated and set to the cutoff position II, the liquid chamber 72 is shut off from the stroke simulator 81 and locked, and the pedal stroke is on the secondary piston 47 ″ side of the master cylinder 2. When the first and second electromagnetic on-off valves 102 and 103 and the electromagnetic switching valve 94 are operated and set to the position II, the pressure conversion cylinder 93 is moved to the power chamber 25. And is connected to the liquid chamber 72, and the liquid chamber 72 is blocked from the stroke simulator 81. At this time, the pedal stroke is equivalent to the stroke of the wheel cylinders 28 and 29. Further, the second electromagnetic on-off valve 103 and the electromagnetic switching valve 94 are both operated and set to position II, is the pressure conversion cylinder 93 cut off from the power chamber 25? It is connected to the liquid chamber 72. At this time, since the liquid chamber 72 is connected to the stroke simulator 81, the pedal stroke is obtained by adding the stroke simulator portion of the stroke simulator 81 and the stroke portions of the wheel cylinders 28 and 29. Become.
[0144]
In this way, by controlling the operation of each electromagnetic valve 94, 102, 103, the pedal stroke can be variously changed. Therefore, a more appropriate pedal stroke can be set depending on the vehicle state such as the loading state, the brake state, or the driver.
[0145]
In the automatic brake device 86, when the automatic brake operation condition is satisfied, the electronic control device not only switches the electromagnetic switching valve 87 to the second position II, but also switches the electromagnetic switching valve 94 to the second position II. Therefore, it is exactly the same as in the case of the fifth example described above, and the automatic brake is activated.
[0146]
Further, in the brake operating device 89 when the hydraulic pressure is lost, when the hydraulic pressure of the accumulator 40 is lost, the electronic control device switches the electromagnetic switching valve 94 to the second position II by a detection signal from the pressure sensor 104. Accordingly, this is exactly the same as in the case of the fifth example described above, and the brake can be operated by depressing the brake pedal even when the hydraulic pressure is lost.
[0147]
FIG. 14 is a diagram partially showing a seventh example of the embodiment of the present invention.
In the fifth example shown in FIG. 11, the first and second switching valves 74 and 75 are used to obtain the jumping characteristic and the reverse two-stage servo characteristic. As shown in FIG. 14, the booster 1 uses third and fourth electromagnetic on-off valves 105 and 106.
[0148]
The third electromagnetic opening / closing valve 105 is provided in a passage connecting the control pressure introduction port 45 and the reservoir 33, and includes a blocking position I that blocks the control pressure introduction port 45 and the reservoir 33, and the control pressure introduction port 45 and the reservoir 33. A communication position II that communicates with 33 is set, and is normally a normally closed on-off valve that is set to a cutoff position I.
[0149]
The fourth electromagnetic opening / closing valve 106 is provided in a passage connecting the control pressure introduction port 45 and the output port 27, and a communication position I for communicating the control pressure introduction port 45 and the output port 27, and the control pressure introduction port 45. And a shutoff position II that shuts off the output port 27 is set, and is normally a normally open on-off valve that is set to the communication position II. Further, the opening and closing of the third and fourth electromagnetic on-off valves 105 and 106 are controlled by the hydraulic pressure in the power chamber 25. For this reason, the hydraulic pressure in the power chamber 25 is detected as described above. A pressure sensor 62 is provided.
Other configurations of the brake hydraulic pressure booster 1 of the seventh example are the same as those of the fifth example.
[0150]
In the brake hydraulic pressure booster 1 of the seventh example configured as described above, when the normal brake operation is performed from the inoperative state shown in FIG. 14, the brake hydraulic pressure booster is performed in exactly the same manner as in the fifth example. The force device 1 performs a jumping action based on the servo ratio of the straight line α ″ shown in FIG. 12. When the hydraulic pressure in the power chamber 25 becomes a pressure corresponding to the end of the jumping action, the electronic device 1 The control device sets the third electromagnetic on-off valve 105 to the communication position I and the fourth electromagnetic on-off valve 106 to the shut-off position II, so that the reaction force chamber 41 is connected to the reservoir 33, so that the brake hydraulic pressure booster The apparatus 1 performs servo control of the normal brake by the servo ratio of the straight line β ″ shown in FIG. When the hydraulic pressure in the power chamber 25 reaches the servo ratio switching point, the electronic control unit moves the third electromagnetic on-off valve 105 to the shut-off position I and the fourth electromagnetic on-off valve 106 based on the detection signal from the pressure sensor 62. Is set to communication position I. Then, since the reaction force chamber 41 is connected to the output port 27, the brake hydraulic pressure booster 1 performs servo control with a large servo ratio of the straight line γ ″ shown in FIG.
Other configurations and operational effects of the brake hydraulic pressure booster 1 of the seventh example are the same as those of the fifth example.
[0151]
In addition, an orifice 100 and a check valve 101 similar to those of the sixth and seventh examples are provided in the pilot pressure introduction passage to the first switching valve 74 of the brake hydraulic pressure booster 1 of the fifth example shown in FIG. Thus, it is possible to obtain a larger braking force in the same manner by delaying the switching of the first switching valve 74 to the second position II during sudden braking.
[0152]
FIG. 15 is a diagram partially showing an eighth example of the embodiment of the present invention.
As shown in FIG. 15, the variable servo device 73 of the brake control device 1 of the eighth example includes the electromagnetic proportional control valve 63 and the pressure adjustment valve 64 shown in FIG. 6, and the accumulator pressure of the accumulator 40 is a pressure. The pressure is adjusted by the adjusting valve 64 and further controlled by the electromagnetic proportional control valve 63 so as to be introduced into the reaction force chamber 41.
Other configurations of the brake hydraulic pressure booster 1 of the eighth example are the same as those of the fifth example.
[0153]
In the brake hydraulic pressure booster 1 of the eighth example, when not in operation, the electromagnetic proportional control valve 63 shuts off the reaction force chamber 41 from the reservoir 33 and connects the pressure regulating valve 64 with the maximum valve opening amount. Therefore, the hydraulic pressure obtained by adjusting the pressure accumulated in the accumulator 40 by the pressure regulating valve 64 is introduced into the reaction force chamber 33. When a normal brake operation is performed in this state, the brake hydraulic pressure booster 1 performs servo control with a large servo ratio of the straight line α ″ shown in FIG. 12, that is, a jumping action.
[0154]
When the hydraulic pressure in the power chamber 25 reaches a level at which the jumping action is finished, the electronic control unit controls the operation of the electromagnetic proportional control valve 63 based on the detection signal of the pressure sensor 62. The hydraulic pressure in the chamber 41 is controlled to a pressure corresponding to the hydraulic pressure in the power chamber 25. As a result, the brake hydraulic pressure booster 1 performs normal servo control with a small servo ratio of the straight line β ″ shown in FIG. 12. When the hydraulic pressure in the power chamber 25 becomes the servo ratio switching point, the electromagnetic proportionality is achieved. The control valve 63 is deactivated again, and the hydraulic pressure regulated by the pressure regulating valve 64 is introduced into the reaction force chamber 41. As a result, the brake hydraulic pressure booster 1 again returns to the straight line α ″ shown in FIG. Servo control is performed with a large servo ratio of the same straight line γ ″.
Thus, also in the brake hydraulic pressure booster 1 of the eighth example, the jumping action and the reverse two-stage servo action are performed.
Other functions and effects of the brake hydraulic pressure booster 1 of the eighth example are the same as those of the fifth example.
[0155]
When shifting from a small servo ratio to a large servo ratio, the electromagnetic proportional control valve 63 is appropriately controlled so that the transition is smooth as shown by the straight lines δ and δ ′ in FIGS. 7 and 10 described above. Can do. Further, as the hydraulic pressure introduced into the reaction force chamber 41, the hydraulic pressure in the power chamber 25 can be used instead of accumulator pressure accumulation.
[0156]
In the fifth to eighth examples described above, the power piston 8 side of the input shaft 18 is braked using the stepped input shaft 18 and the reaction force piston 20 having a small diameter on the power piston 8 side and a large diameter on the brake pedal side. A step portion having a larger diameter than the pedal side is formed, and this step portion is positioned in the reaction force chamber 41. However, in each of the following examples, the reaction force piston 20 is omitted, and the power piston 8 side has a small diameter and a brake. The stepped portion of the stepped input shaft 18 having a large diameter on the pedal side is positioned in the reaction force chamber 41.
[0157]
FIG. 16 is a diagram partially showing a ninth example of the embodiment of the present invention.
In the fifth example shown in FIG. 11, the reaction force piston 20 causes the outer diameter of the input shaft 18 on the power piston 8 side to be substantially equal to the outer diameter of the reaction force piston 20 (pressure receiving area A). _Three Therefore, the outer diameter of the brake pedal (pressure receiving area A ₂ The hydraulic pressure in the reaction force chamber 41 acts on the step portion of the outer diameter of the reaction force piston 20 and the outer diameter of the input shaft 18 on the brake pedal side in the same direction as the input of the input shaft 18. However, as shown in FIG. 16, in the brake hydraulic pressure booster 1 of the ninth example, the reaction force piston 20 is omitted, and therefore the outer diameter of the input shaft 18 on the power piston 8 side. (Pressure receiving area A ₁ ) Is the outer diameter of the brake pedal (pressure receiving area A) ₂ The hydraulic pressure in the reaction force chamber 41 is applied to the step portion 18a of the input shaft 18 in the opposite direction to the input of the input shaft 18.
[0158]
The variable servo device 73 connected to the control pressure introduction port 45 is arranged in series with the first switching valve 107 for performing the jumping operation and the first switching valve 107, and performs the reverse two-stage servo operation. It is comprised from the 2nd switching valve 108 for performing.
[0159]
The first switching valve 107 is a two-position three-way valve, and is controlled by a pilot pressure based on the hydraulic pressure in the power chamber 25. In the first switching valve 107, a first position I for connecting the second switching valve 108 to the reservoir 33 and a second position II for connecting the second switching valve 108 to the output port 27 are set. When the brake hydraulic pressure booster 1 is not in operation, it is set to the first position I, and when the hydraulic pressure in the power chamber 25 reaches the first set operating pressure for terminating the jumping action, it switches to the second position II. It is set up.
[0160]
Similarly, the second switching valve 108 is a two-position three-way valve, and is controlled by a pilot pressure generated by the hydraulic pressure in the power chamber 25. In the second switching valve 108, a first position I for connecting the control pressure introduction port 45 to the first switching valve 107 and a second position II for connecting the control pressure introduction port 45 to the reservoir 33 are set. Normally, the first position I is set, and when the hydraulic pressure in the power chamber 25 reaches the second set operating pressure set to the hydraulic pressure at the servo ratio switching point, the first position I is switched to the second position II. It has become so.
Other configurations of the brake hydraulic pressure booster 1 of the ninth example are the same as those of the fifth example.
[0161]
In the brake hydraulic pressure booster 1 of the ninth example configured as described above, the brake hydraulic pressure booster 1 is in the state shown in the figure when not operating. In this state, when the hydraulic pressure is introduced into the power chamber 25 by the normal brake operation, the power piston 8 moves forward. At this time, the hydraulic pressure in the power chamber 25 is not introduced into the reaction chamber 41 because the first switching valve 107 is set at the first position I. As a result, the brake fluid pressure booster 1 performs servo control with a servo ratio larger than the servo ratio during normal braking, as in the above example. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are given by the above-described mathematical formula 1 and also represented by a straight line α ″ as shown in FIG. 12. Therefore, the brake hydraulic pressure booster 1 is The jumping action is performed by the servo ratio of the straight line α ″.
[0162]
When the hydraulic pressure in the power chamber 25 reaches the first set operating pressure, the first switching valve 107 is operated to switch to the second position II, the reaction force chamber 41 is connected to the output port 27, and the liquid in the power chamber 25 is Pressure is introduced into the reaction force chamber 41. As a result, the servo ratio is reduced to be the same as the servo ratio during normal braking, and thereafter, the brake hydraulic pressure booster 1 performs servo control with the servo ratio during normal braking. The input / output characteristic of the brake hydraulic pressure booster 1 at this time is given by Expression 6 and is represented by a straight line β ″ as shown in FIG.
[0163]
When the hydraulic pressure in the power chamber 25 reaches the second set operating pressure, the second switching valve 108 is operated and switched to the second position II, and the reaction force chamber 41 is connected to the reservoir 33. Then, the pressure liquid in the reaction force chamber 41 is discharged to the reservoir 33, and the reaction force chamber 41 becomes atmospheric pressure. For this reason, servo control is performed with a servo ratio that is greater than the servo ratio during normal braking. The input / output characteristics of the brake hydraulic pressure booster 1 at this time are given by the above-described mathematical formula 1 and also represented by a straight line γ ″ that is the same straight line as the straight line α ″ as shown in FIG.
[0164]
Thus, in the brake hydraulic pressure booster 1 of the ninth example, a jumping action is performed by servo control with a large servo ratio along the straight line α ″ at the initial stage of brake operation, and then along the straight line β ″. When a normal braking action is performed by a servo control with a small servo ratio and the input is increased to some extent, a brake control requiring a large braking force is performed by a servo control with a large servo ratio along the straight line γ ″. Thus, even if the brake hydraulic pressure booster 1 of the ninth example does not have the reaction force piston 20, it has a jumping characteristic due to the first switching valve 107, and the reverse switching two-stage due to the second switching valve 108. Has servo characteristics.
Other functions and effects of the ninth example of the brake hydraulic pressure booster 1 are the same as those of the fifth example.
[0165]
FIG. 17 is a diagram partially showing a tenth example of the embodiment of the present invention.
In the ninth example shown in FIG. 16, the supply and discharge of the pressure fluid to and from the reaction force chamber 41 is controlled by the two switching valves 107 and 108. In the variable servo device 73 of the brake control device 1 of the tenth example, As shown in FIG. 17, an electromagnetic switching valve 109 is provided in place of these switching valves 107 and 108. The variable servo device 73 includes a piston 96, a cylinder 97, and a spring 98, and a low-pressure accumulator 99 for accumulating liquid, a passage between the indicated pressure introduction port 45, the electromagnetic switching valve 109, and the low-pressure accumulator 99. And a check valve 101 that bypasses the orifice 100 and permits only the flow of liquid from the indicated pressure introduction port 45 to the electromagnetic switching valve 109 and the low-pressure accumulator 99.
[0166]
The electromagnetic switching valve 109 is set to a first position I where the control pressure introduction port 45 is connected to the reservoir 33 during non-operation and a second position II where the control pressure introduction port 45 is connected to the output port 27 during operation. . Then, the electronic control unit moves the electromagnetic switching valve 109 to the inoperative first position I until the hydraulic pressure in the power chamber 25 detected by the pressure sensor 62 reaches the jumping end pressure at which the conventional jumping action ends. When the hydraulic pressure in the power chamber 25 reaches the jumping action end pressure, the electromagnetic switching valve 109 is operated to set the second position II, and the hydraulic pressure in the power chamber 25 is set to the pressure at the servo ratio switching point. Then, the electromagnetic switching valve 109 is set to the non-actuated first position I again.
Other configurations of the brake hydraulic pressure booster 1 of the tenth example are the same as those of the ninth example.
[0167]
In the brake hydraulic pressure booster 1 of the tenth example, when the brake is not operated, the reaction force chamber 41 is connected to the reservoir 33 and is at atmospheric pressure. In this state, when the normal brake operation is performed The brake hydraulic pressure booster 1 performs servo control with a large servo ratio of the straight line α ″ shown in FIG. 12, that is, a jumping action.
[0168]
When the hydraulic pressure in the power chamber 25 reaches a level at which the jumping action is finished, the electronic control unit switches the electromagnetic switching valve 109 to the second position II based on the detection signal of the pressure sensor 62, and the reaction force chamber 41 is switched to the power chamber 41. 25 pressure fluids are introduced. At this time, since it is during normal braking, the pressure fluid flows at a normal speed and is not throttled by the orifice 100. As a result, the brake hydraulic pressure booster 1 completes the jumping action without delay, and performs normal servo control with a small servo ratio of the straight line β ″ shown in FIG. 12. The hydraulic pressure in the power chamber 25 is When the servo ratio switching point is reached, the electromagnetic switching valve 109 is switched again to the first position I, and the reaction force chamber 41 is again connected to the reservoir 33. Therefore, the pressure fluid in the power chamber 25 is discharged to the reservoir 33, and the reaction force The hydraulic pressure in the chamber 41 becomes atmospheric pressure, whereby the brake hydraulic pressure booster 1 again performs servo control with a large servo ratio of the same straight line γ ″ as the straight line α ″ shown in FIG.
[0169]
By the way, when a sudden braking operation is performed, the hydraulic pressure in the power chamber 25 rapidly increases. Then, when the electromagnetic switching valve 109 is switched to the second position II, the pressure fluid in the power chamber 25 rapidly flows to the reaction force chamber 41. This pressure fluid is throttled by the orifice 100 and is also a low-pressure accumulator. 99. As a result, the low-pressure accumulator 99 is accumulated. For this reason, since the increase in the hydraulic pressure in the reaction force chamber 41 is delayed, the shift from the large servo ratio of the straight line α ″ shown in FIG. 12 to the small servo ratio of the straight line β ″ is delayed. That is, the servo ratio switching point changes and the end of the jumping action is delayed. Therefore, the brake hydraulic pressure booster 1 accordingly generates a large output, and can obtain a larger braking force during sudden braking.
[0170]
Thus, the brake hydraulic pressure booster 1 of the tenth example also performs the jumping action and the reverse two-stage servo action.
Other effects of the brake hydraulic pressure booster 1 of the tenth example are the same as those of the ninth example.
[0171]
FIG. 18 is a diagram partially showing an eleventh example of the embodiment of the present invention.
As shown in FIG. 18, in the variable servo device 73 of the brake control device 1 of the eleventh example, two third and fourth electromagnetic on-off valves 110, 110 are provided as in the seventh example shown in FIG. ing.
[0172]
The third electromagnetic on-off valve 110 is the same normally open on-off valve as the fourth electromagnetic on-off valve 106 of the seventh example, and the fourth electromagnetic on-off valve 111 is the same normally closed as the third electromagnetic on-off valve 105 of the seventh example. Open / close valve.
Other configurations of the brake hydraulic pressure booster 1 of the eleventh example are the same as those of the ninth example.
[0173]
In the brake hydraulic pressure booster 1 of the eleventh example, when the brake is not operated, the reaction force chamber 41 is connected to the reservoir 33 and is at atmospheric pressure. When the normal brake operation is performed in this state, The brake hydraulic pressure booster 1 performs servo control with a large servo ratio of the straight line α ″ shown in FIG. 12, that is, a jumping action.
[0174]
When the hydraulic pressure in the power chamber 25 is large enough to complete the jumping action, the electronic control unit switches both the third and fourth electromagnetic switching valves 110 and 111 to the second position II based on the detection signal of the pressure sensor 62. The pressure fluid from the power chamber 25 is introduced into the reaction force chamber 41. As a result, the brake hydraulic pressure booster 1 performs normal servo control with a small servo ratio of the straight line β ″ shown in FIG. 12. When the hydraulic pressure in the power chamber 25 becomes the servo ratio switching point, the third control is performed. And the fourth electromagnetic switching valves 110 and 111 are both switched again to the first position I, and the reaction force chamber 41 is connected to the reservoir 33 again, so that the pressure fluid in the power chamber 25 is discharged to the reservoir 33 and the reaction force chamber. The hydraulic pressure at 41 becomes the atmospheric pressure, whereby the brake hydraulic pressure booster 1 again performs servo control with a large servo ratio of the same straight line γ ″ as the straight line α ″ shown in FIG.
[0175]
Thus, also in the brake hydraulic pressure booster 1 of the eleventh example, the jumping action and the reverse two-stage servo action are performed.
Other functions and effects of the brake hydraulic pressure booster 1 of the eleventh example are the same as those of the ninth example.
[0176]
FIG. 19 is a diagram partially showing a twelfth example of the embodiment of the present invention.
As shown in FIG. 19, in the variable servo device 73 of the brake control device 1 of the twelfth example, the electromagnetic proportional control valve 63 and the pressure adjustment valve 64 shown in FIG. 6 are provided, and the accumulator pressure of the accumulator 40 is a pressure. The pressure is adjusted by the adjusting valve 64 and further controlled by the electromagnetic proportional control valve 63 so as to be introduced into the reaction force chamber 41.
Other configurations of the brake hydraulic pressure booster 1 of the twelfth example are the same as those of the ninth example.
[0177]
In the brake hydraulic pressure booster 1 of the twelfth example, when not operating, the electromagnetic proportional control valve 63 connects the reaction force chamber 41 to the reservoir 33, and the reaction force chamber 41 is at atmospheric pressure. When a normal brake operation is performed in this state, the brake hydraulic pressure booster 1 performs servo control with a large servo ratio of the straight line α ″ shown in FIG. 12, that is, a jumping action.
[0178]
When the hydraulic pressure in the power chamber 25 reaches a level at which the jumping action is finished, the electronic control unit controls the operation of the electromagnetic proportional control valve 63 based on the detection signal of the pressure sensor 62. A pressure corresponding to the hydraulic pressure of 25 is output, and the pressurized liquid of this output pressure is introduced into the reaction force chamber 41. As a result, the brake hydraulic pressure booster 1 performs normal servo control with a small servo ratio of the straight line β ″ shown in FIG. 12. When the hydraulic pressure in the power chamber 25 becomes the servo ratio switching point, the electromagnetic proportionality is achieved. Since the control valve 63 is deactivated and the reaction force chamber 41 is connected to the reservoir 33 again, the pressure fluid in the power chamber 25 is discharged to the reservoir 33, and the fluid pressure in the reaction force chamber 41 becomes atmospheric pressure. The brake hydraulic pressure booster 1 again performs servo control with a large servo ratio of the same straight line γ ″ as the straight line α ″ shown in FIG.
[0179]
As described above, the brake hydraulic pressure booster 1 of the twelfth example also performs the jumping action and the reverse two-stage servo action.
Other effects of the brake hydraulic pressure booster 1 of the twelfth example are the same as those of the ninth example.
[0180]
When shifting from a small servo ratio to a large servo ratio, the electromagnetic proportional control valve 63 is appropriately controlled so that the transition is smooth as shown by the straight lines δ and δ ′ in FIGS. 7 and 10 described above. Can do. Further, as the hydraulic pressure introduced into the reaction force chamber 41, the hydraulic pressure in the power chamber 25 can be used instead of accumulator pressure accumulation.
In the above-described embodiment, the hydraulic booster of the present invention is described as being applied to a brake hydraulic booster. However, the hydraulic booster can be applied to other hydraulic boosters other than brakes.
[0181]
【The invention's effect】
As is clear from the above description, according to the hydraulic booster of the present invention, it is possible to obtain reverse two-stage servo characteristics and a jumping characteristic with a simple structure. As a result, an output larger than a normal output can be obtained with a predetermined input or more.
[0182]
In addition, the servo ratio switching point can be changed, so that various input / output characteristics can be flexibly handled.
Furthermore, since the electromagnetic proportional control valve is used, the servo ratio can be changed smoothly.
[0183]
Furthermore, according to the brake control system of the present invention, the rise of the braking force can be accelerated by the jumping characteristic with a simple structure, and a large braking force can be obtained quickly during sudden braking by the reverse two-stage servo control. Even a driver such as a beginner can assist to ensure that a large braking force can be obtained.
[Brief description of the drawings]
FIG. 1 is a cross-sectional view showing a first example of an embodiment of a brake hydraulic pressure booster according to the present invention.
FIG. 2 is a partially enlarged cross-sectional view of the brake hydraulic pressure booster shown in FIG.
3 is a cross-sectional view showing a reaction force piston used in the brake fluid pressure booster shown in FIG. 1. FIG.
FIG. 4 is a diagram showing input / output characteristics of the brake hydraulic pressure booster shown in FIG. 1;
FIG. 5 is a sectional view showing a second example of the embodiment of the present invention.
FIG. 6 is a cross-sectional view showing a third example of the embodiment of the present invention.
7 is a graph showing input / output characteristics of the brake fluid pressure booster shown in FIG. 6;
FIG. 8 is a sectional view showing a fourth example of the embodiment of the present invention.
9 is a diagram showing input / output characteristics of the brake fluid pressure booster shown in FIG.
10 is a diagram showing input / output characteristics of a modified example of the brake hydraulic pressure booster shown in FIG.
FIG. 11 is a cross-sectional view showing a fifth example of the embodiment of the present invention.
FIG. 12 is a diagram showing input / output characteristics of fifth to eleventh examples of the embodiment of the present invention.
FIG. 13 is a sectional view showing a sixth example of the embodiment of the present invention.
FIG. 14 is a cross-sectional view showing a seventh example of the embodiment of the invention.
FIG. 15 is a sectional view showing an eighth example of the embodiment of the present invention.
FIG. 16 is a sectional view showing a ninth example of the embodiment of the present invention;
FIG. 17 is a sectional view showing a tenth example of the embodiment of the present invention.
FIG. 18 is a sectional view showing an eleventh example of the embodiment of the present invention.
FIG. 19 is a sectional view showing a twelfth example of the embodiment of the present invention.
FIG. 20 is a partial sectional view partially showing a conventional brake hydraulic pressure booster.
FIG. 21 is a diagram showing input / output characteristics of the brake hydraulic pressure booster shown in FIG. 20;
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 1 ... Brake fluid pressure booster, 2 ... Master cylinder, 3 ... Housing, 8 ... Power piston, 18 ... Input shaft, 18a ... Step part, 20 ... Reaction force piston, 20e ... Right end of reaction force piston, 21 ... Spring , 25 ... power chamber, 27 ... output port, 33 ... reservoir, 40 ... accumulator, 41 ... reaction force chamber, 44 ... master cylinder piston, 45 ... control pressure inlet, 46 ... switching valve, 60 ... control valve, 61, 95, 109 ... electromagnetic switching valve, 62 ... pressure sensor, 63 ... electromagnetic proportional control valve, 64 ... pressure regulating valve, 73 ... variable servo device, 74, 107 ... first switching valve, 75, 108 ... second switching valve, 99 ... Low pressure accumulator, 100 ... Orifice, 101 ... Check valve, 105 ... First electromagnetic on-off valve, 106 ... Second electromagnetic on-off valve, 110 ... Third electromagnetic switching valve, 111 ... Fourth electromagnetic switching valve

Claims (14)

A hydraulic pressure source for generating hydraulic pressure, a reservoir for storing hydraulic fluid, a power piston for generating output, a power chamber facing the pressure receiving surface of the power piston, and the power chamber from the hydraulic pressure source when not operating Control that shuts off and communicates with the reservoir, shuts off the power chamber from the reservoir during operation and communicates with the hydraulic pressure source, and introduces the hydraulic fluid from the hydraulic pressure source into the power chamber according to the actuation A valve, a stepped input shaft having a small diameter on the control valve side and a large diameter step on the side opposite to the control valve, and an outer diameter of the input shaft. Is set to be larger than the diameter of the portion on the opposite side of the control valve and is formed in a cylindrical shape and is slidably fitted to the small diameter portion of the input shaft, with its front end facing the power chamber and thereafter Reaction force fixie whose end can contact the step of the input shaft And the rear end of the reaction force piston is always urged away from the stepped portion of the input shaft, and the reaction force piston rear end is input to the input when the hydraulic pressure in the power chamber is equal to or higher than a first predetermined pressure. A spring to be brought into contact with a step portion of the shaft, and an operation means for operating the input shaft when an input is applied,
A reaction force chamber is formed so that the step portion of the input shaft is located, and the servo ratio control pressure is set in the reaction force chamber when the hydraulic pressure of the power chamber is equal to or higher than a second predetermined pressure higher than the first predetermined pressure. And a servo ratio control pressure control valve is provided for controlling so that the pressure fluid in the reaction force chamber is discharged to the reservoir when the fluid pressure in the power chamber is smaller than the second predetermined pressure. A hydraulic booster characterized by comprising: A hydraulic pressure source for generating hydraulic pressure, a reservoir for storing hydraulic fluid, a power piston for generating output, a power chamber facing the pressure receiving surface of the power piston, and the power chamber from the hydraulic pressure source when not operating Control that shuts off and communicates with the reservoir, shuts off the power chamber from the reservoir during operation and communicates with the hydraulic pressure source, and introduces the hydraulic fluid from the hydraulic pressure source into the power chamber according to the actuation A valve, a stepped input shaft having a control valve side portion having a large diameter and a portion on the opposite side of the control valve side having a small diameter, and an input applied to the control valve. Operating means for operating the input shaft,
A reaction force chamber is formed so that the stepped portion of the input shaft is positioned, and when the hydraulic pressure of the power chamber is equal to or higher than a predetermined pressure, the pressure force of the servo ratio control pressure is introduced into the reaction force chamber, and A hydraulic pressure booster comprising a servo ratio control pressure control valve for controlling the hydraulic fluid in the reaction force chamber to be discharged to the reservoir when the hydraulic pressure in the power chamber is smaller than the predetermined pressure. . The hydraulic booster according to claim 1 or 2, wherein the servo ratio control pressure control valve comprises one switching valve or two on-off valves. The operation of the switching valve or the on-off valve is controlled by a hydraulic pressure of the power chamber or an electromagnetic force excited according to the hydraulic pressure of the power chamber. Hydraulic booster. A hydraulic pressure source for generating hydraulic pressure, a reservoir for storing hydraulic fluid, a power piston for generating output, a power chamber facing the pressure receiving surface of the power piston, and the power chamber from the hydraulic pressure source when not operating Control that shuts off and communicates with the reservoir, shuts off the power chamber from the reservoir during operation and communicates with the hydraulic pressure source, and introduces the hydraulic fluid from the hydraulic pressure source into the power chamber according to the actuation A valve, a stepped input shaft having a control valve side portion having a large diameter and a portion on the opposite side of the control valve side having a small diameter, and an input applied to the control valve. Operating means for operating the input shaft,
The stepped portion of the input shaft is formed with a reaction chamber so as to be positioned, the reaction force first plant hydraulic pressure before Symbol power chamber to the chamber pressure is less than the time and the first predetermined pressure is greater than the second When the pressure is higher than a predetermined pressure, the hydraulic pressure of the servo ratio control pressure is introduced, and when the hydraulic pressure in the power chamber is higher than the first predetermined pressure and lower than the second predetermined pressure, the pressure fluid in the reaction force chamber is discharged to the reservoir. A hydraulic pressure booster is provided with a servo ratio control pressure control valve for controlling the pressure to be controlled. An orifice is provided in a passage for introducing hydraulic pressure in the power chamber for obtaining an operating pressure signal for switching and controlling the servo ratio control pressure control valve by the hydraulic pressure in the power chamber. The hydraulic booster according to claim 5. A hydraulic pressure source for generating hydraulic pressure, a reservoir for storing hydraulic fluid, a power piston for generating output, a power chamber facing the pressure receiving surface of the power piston, and the power chamber from the hydraulic pressure source when not operating Control that shuts off and communicates with the reservoir, shuts off the power chamber from the reservoir during operation and communicates with the hydraulic pressure source, and introduces the hydraulic fluid from the hydraulic pressure source into the power chamber according to the actuation A valve, a stepped input shaft having a step portion with a small diameter on the control valve side and a large diameter portion on the opposite side to the control valve side, and an input applied to the control valve. Operating means for operating the input shaft,
And the reaction force chamber is formed as a stepped portion of the input shaft is located, the hydraulic pressure of the front Symbol power chamber to the reaction chamber is and the first predetermined pressure or more the first predetermined pressure is greater than the second predetermined When the pressure is smaller than the pressure, the pressure ratio of the servo ratio control pressure is introduced, and when the pressure in the power chamber is smaller than the first predetermined pressure and the pressure in the reaction chamber is higher than the second predetermined pressure, A hydraulic pressure booster is provided with a servo ratio control pressure control valve that controls the exhaust to be discharged. The passage between the servo ratio control pressure control valve and the reaction force chamber is provided with at least an orifice and a low pressure accumulator located on the servo ratio control pressure control valve side from the orifice. The hydraulic booster according to claim 7. The hydraulic pressure according to any one of claims 5 to 8, wherein the servo ratio control pressure control valve includes two switching valves, and these switching valves are controlled by hydraulic pressure in the power chamber. Boost device. The servo ratio control pressure control valve is composed of a switching valve or two on-off valves, and the switching valve or the on-off valve is controlled by an electromagnetic force excited in accordance with a hydraulic pressure in the power chamber. The hydraulic booster according to any one of claims 5 to 8. The hydraulic booster according to any one of claims 1 to 10, wherein the servo ratio control pressure is a hydraulic pressure of the power chamber. The servo ratio control pressure control valve is an electromagnetic proportional control valve that is controlled according to the hydraulic pressure of the power chamber, and the servo ratio control pressure is the hydraulic pressure of the power chamber or the hydraulic pressure of the hydraulic pressure source. 8. The hydraulic booster according to claim 1, wherein the hydraulic pressure is controlled by the electromagnetic proportional control valve. The hydraulic booster according to any one of claims 1 to 12, a master cylinder that is controlled by the output of the hydraulic booster to generate brake hydraulic pressure, and the brake hydraulic pressure of the master cylinder is introduced. And a brake cylinder for generating a braking force by being applied. In the two brake systems,
The hydraulic booster according to any one of claims 1 to 12, a master cylinder that is controlled by the output of the hydraulic booster to generate a brake hydraulic pressure, and the power of the hydraulic booster A brake cylinder of one system that generates a braking force by introducing the hydraulic pressure of the chamber, and a brake cylinder of the other system that generates a braking force by introducing the brake hydraulic pressure of the master cylinder. The brake hydraulic pressure booster is characterized by a semi-full power brake.

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