JP3564911B2

JP3564911B2 - Hydraulic drive

Info

Publication number: JP3564911B2
Application number: JP01007697A
Authority: JP
Inventors: 智浜本
Original assignee: Nachi Fujikoshi Corp
Current assignee: Nachi Fujikoshi Corp
Priority date: 1996-01-08
Filing date: 1997-01-06
Publication date: 2004-09-15
Anticipated expiration: 2017-01-06
Also published as: JPH1089304A

Description

【０００１】
【発明の属する技術分野】
本発明は建設機械等で使用される１つの油圧ポンプの吐出油を複数のアクチュエータのそれぞれに流入する圧油を制御可能にされた流量調節機能を有する複数の方向制御弁及び各方向制御弁のそれぞれの圧力補償をする複数の圧力補償弁とを有する油圧駆動装置に関する。
【０００２】
【従来の技術】
この種の油圧駆動装置は建設機械や農業機械用に主として用いられ、負荷圧力に応じて可変ポンプ吐出量を制御するロードセンシング機能を備えたものが使用されている。また、複数のアクチュエータを同時に駆動するとき、各アクチュエータの負荷圧力の差により互いに干渉してアクチュエータの速度変化を生じないように、各アクチュエータへの回路に圧力補償弁を設けることにより、ポンプ吐出量を分流するようにされている。さらに、ポンプ吐出量が複数の駆動アクチュエータの所定要求流量を下まわった場合には、各アクチュエータに適切な比でポンプ吐出量を分配させる機能いわゆるアンチサチュレーション機能を備えたものも使用されている。例えば、かかるアンチサチュレーション機能を備えた特公平４−４８９６７号公報のものにおいては、ロードセンシング機能はスプリング力と各アクチュエータのうちの最高負荷圧力とをそれぞれ可変ポンプ吐出量を増加させる方向に作用させ、この作用力に対向して可変ポンプ吐出量を減少させる方向に吐出圧力を作用させるようしたポンプ流量制御部を設け、負荷圧力に応じてポンプ吐出量を制御する。一方、例えば特開平４─１９４０９号公報に記載するように、流量調整機能を有する方向制御弁の下流に圧力補償弁を配置し、圧力補償弁の各制御油室に圧力補償弁を開く方向に圧力補償弁の上流側の圧力（Ｐｄ′）をそれぞれ作用させ、圧力補償弁を絞る方向にアクチュエータの最高負荷圧力（Ｐｍ）をそれぞれ作用させるようにしてアンチサチュレーション機能をもたせるものが開示されている。
また実開平４−１１９６０４号公報では、圧力補償弁の各制御油室に、圧力補償弁を閉じる方向に圧力補償弁の下流側の圧力（Ｐｚ）を作用させ、可変ポンプ吐出圧力（Ｐｄ）とアクチュエータの最高負荷圧力（Ｐｍ）との差圧に対応した二次圧（Ｐｃ＝Ｐｄ−Ｐｍ）を発生する差圧制御弁を設け、圧力補償弁を開く方向にこの差圧制御弁から出力される二次圧力（Ｐｃ）及び方向制御弁の下流側圧力であるアクチュエータ負荷圧力（ＰＬ）を作用させるようにしたものも記載されている。さらに特開平４−５４３０３号公報第６図には、可変ポンプ吐出圧力（Ｐｄ）とアクチュエータの最高負荷圧力（Ｐｍ）を検知する差圧検出器と、差圧検出器の出力を受けて制御信号を出力する制御装置と、制御装置の制御信号で作動する電磁比例弁が出力する二次圧力（Ｐｃ）によって、アンチサチュレーション機能を確保するものも開示されている。
【０００３】
このような従来のアンチサチュレーション機能を有する圧力補償弁の流量補償域の流量特性は、開き側と閉じ側との釣り合いにより、アクチュエータ負荷圧力（ＰＬ）と方向制御弁の上流側圧力（Ｐｚ）との差圧、即ち、方向制御弁の前後の差圧（ΔＰ：以下方向制御弁差圧という）の関係は、
ΔＰ＝Ｐｚ−ＰＬ＝Ｐｄ−Ｐｍ＝Ｐｃ
で示され、方向制御弁差圧が可変ポンプ吐出圧力とアクチュエータの最高負荷圧力との差圧、即ち二次圧力、に比例した圧力にされている。
しかし、かかる圧力補償を行っている油圧駆動装置において、慣性の大きな負荷を動かした場合には、系が安定せず、ハンチングを伴なった作動をすることが知られている。例えば、建設機械の油圧ショベルを動かす場合に、負荷の大きい旋回モータ、走行モータ、又はブームシリンダのように負荷の大きいシリンダを動かす場合などでは、ハンチングが顕著に表われ、操作性が損なわれるという問題があった。即ち、今、方向制御弁のレバーを一定量だけステップ的に動かし、慣性の大きいアクチュエータすなわち旋回モータ等を動かす場合を考える。まず、方向制御弁の絞りが開かれ、油が流れ込むがアクチュエータの慣性が大きいため、アクチュエータは急には動かないので流量は流れずに負荷圧力が一瞬にして上昇する。負荷圧力が上昇すると、その負荷圧力が圧力補償弁に作用し、圧力補償弁を大きく開く方向に働く。そのため大流量が供給されアクチュエータは急激に加速されることになるが、一旦動き出すと供給される流量は有限であるため、速度は増していくが、加速度は次第に小さくなる。従って急上昇した負荷圧力は加速度の減少に伴ない徐々に小さくなっていくので、圧力補償弁の開度が徐々に小さくなっていき、供給される流量が減少する。一方、アクチュエータは加速度がなくなり一定の速度となると、その速度は初期の大きな加速度により加速されたものであるから、目標とする速度よりもかなり大きいと同時に、その時点での負荷圧力は加速がおさまった後であることからかなり小さくなっているので、今度は圧力補償弁の開度はより小さくなり、それゆえ方向制御弁差圧も小さくなっている。このため流量が減少しアクチュエータが減速し始めることになるが、大きな慣性のため、アクチュエータはその速度を保とうとすることにより、負荷圧力はさらに急激に減少することになる。そこで圧力補償弁の開度は増々小さくなり、アクチュエータの速度はより減少するが、速度がある程度小さくなると逆に、負荷圧力は徐々に回復し、それ故に圧力補償弁の開度は徐々に大きくなっていき、そのうちにアクチュエータの減速は止まり一定の速度となるが、その速度は減速の初期に急激に減少したため、目標とする速度よりかなり小さいと同時に、その時点での負荷圧力は、減速がおさまった後であることから、大きな値に回復しているので圧力補償弁の開度も大きくなっており、それゆえ方向制御弁差圧も大きくなっていることから、今度はアクチュエータが加速し始めることになる。
【０００４】
一旦アクチュエータが加速し始めると前述した最初の状態となる。こうして、急加速と急減速を繰り返す現象はなかなか収束せずハンチングが継続する。実際には、さらに、ポンプ装置側の応答遅れが加わり、もっと複雑な様相を呈する。このように圧力補償を行っている回路においては、慣性の大きな負荷を動かした場合に、系が安定せず、ハンチングが生じ易いという問題があった。しかしながら前述した、特公平４−４８９６７号公報や実開平４−１１９６０４号公報のものは、圧力補償弁の各制御室に作用する開閉方向の作用圧力が相等しくされており、また、アンチサチュレーション機能を有することが開示されているが、ハンチングの防止については開示も示唆もされていない。特開平４−５４３０３号公報では、複数のアクチュエータの同時作動時に、圧力補償弁の絞り部によって発生するフローフォースによる流量減少によって生じる圧力補償弁のハンチングや不安定現象をごくわずかな受圧面積でメインポンプの吐出圧力を圧力補償弁が開く方向に作用させ、ポンプの吐出圧力とアクチュエータの負荷圧力との差圧が増加すると、圧力補償弁の出口流量を増加させて、フローフォースをキャンセルしてフローフォースに影響されない出口流量を得るようにしたものが開示されているが、慣性の大きい負荷に対するハンチングの防止については開示も示唆もされていない。
さらに、上記従来技術ではいずれも、最高負荷圧力（Ｐｍ）をバルブ装置から長く細いパイロット油路を介してポンプの押しのけ容積を変更するポンプ流量調整弁を閉じる方向に作用させていた。このため低温時に作動油の粘度が高くなり、ポンプからバルブ装置までの管路での圧力損失が過大なものとなるとバルブ装置内の圧力補償弁の上流側の圧力が前記の圧力損失分だけ低下し、そのため方向制御弁差圧が低下してアクチュエータへの供給流量が大幅に減少した。
また、油圧走行車両の一対のクローラを回転駆動させる２個の走行モータ駆動のように、複数のアクチュエータのうちの少なくとも２個のアクチュエータがアクチュエータの負荷圧力に拘わらず同期して駆動する必要のあるときに、各方向制御弁を同じストロークだけ切換操作すれば、圧力補償弁により各方向制御弁前後の方向制御弁差圧が等しくなるよう制御されるので各走行モータには等しい流量が供給されることになり油圧走行車両は直進走行できるはずであるが、各方向制御弁のスプールの加工誤差があると方向制御弁差圧を等しくしても夫々の方向制御弁の絞り開度が異なることになるので各走行モータに供給される流量は等しくならず、また圧力補償弁の加工誤差による受圧面積の誤差があると、方向制御弁の絞り開度が等しくても、方向制御弁差圧が等しくならないので、油圧走行車両の直進走行が出来なくなるという問題があった。
【０００５】
さらにまた、油圧ショベルの旋回用油圧モータと油圧ブームシリンダーのように、負荷が大きく異なる少なくとも２個の油圧アクチュエータを同時に作動させるとき、同時操作の初期には高負荷側アクチュエータの慣性負荷が過大であるため流入側のアクチュエータポートに過大な圧力が発生し、圧油のほとんどが流入側アクチュエータポートに設置されているオーバロードリリーフ弁よりタンクへ流出し、このため有効吐出流量そのものが減少し、低負荷側油圧アクチュエータであるブームシリンダの駆動速度が極端に遅くなると同時に旋回モータへ流入した圧油のリリーフによる原動機のエネルギー損失が大きくなるという問題があった。その後旋回モータの加速が終了し、定常速度による旋回となると、旋回モータの負荷圧力が急激に低下する。旋回モータ用の圧力補償弁は初期の段階で過大な旋回負荷圧力により、ほぼ全開となっていたが、負荷圧力が急激に低下することでその開度は急激に小さくなる。そのため旋回モータはショックを伴って減速する。その減速に伴いポンプ（有効）吐出量に余裕が出ることにより、ブームの方は逆に加速し、ギクシャクした動きとなるという問題もあった。
【０００６】
【発明が解決しようとする課題】
本発明のうち請求項１ないし１４に記載の発明の目的は、従来のかかる問題点を鑑みなされたもので、その目的は、単独操作、複合操作であっても、低負荷側、高負荷側共操作性の良いハンチングの生じない安定した圧力補償弁を有する油圧駆動装置を提供することである。また、単純な構造とし、コスト安で、高い信頼性を得られ、さらには、負荷条件に応じて柔軟に対応できる圧力補償弁を有する油圧駆動装置を提供することにある。本発明のうち請求項１記載の発明の目的は、低温時に作動油の粘度が高くなり、ポンプからバルブ装置までの管路での圧力損失が過大なものとなり、アクチュエータへの供給流量が大幅に減少することを防止する油圧駆動装置を提供することである。
本発明のうち請求項５及び 12 記載の発明の目的は、油圧走行車両の一対のクローラを回転駆動させる２個の走行モータでは、複数のアクチュエータのうちの少なくとも２個のアクチュエータがアクチュエータの負荷圧力に拘わらず同期して駆動する必要のあるため、各方向制御弁のスプールの加工誤差、圧力補償弁の加工誤差による受圧面積の誤差があるときにおいても、油圧走行車両の直進走行が得ることができる圧力補償弁を有する油圧制御装置を提供することにある。
本発明のうち請求項６、７、 13 及び 14 に記載の発明の目的は、極端に負荷の大きさが異なったアクチュエータを同時に操作しても、小さな負荷側へも、充分に圧油を供給することができ、かつ、極めて大きい負荷特性のアクチュエータの負荷圧力が急激に低下した場合でも、夫々のアクチュエータの速度が急変せずにショックなくスムースな操作を可能とし、かつエネルギー損失や原動機の負担を減少することができる圧力補償弁を有する油圧駆動装置を提供することにある。
【０００７】
【課題を解決するための手段】
このために、本発明の第１発明によると、請求項２、１０記載の発明のように、可変ポンプと、該可変ポンプの吐出油によって駆動される複数の油圧アクチュエータと、該複数のアクチュエータのそれぞれに流入する圧油を制御可能にされた流量調節機能を有する複数の方向制御弁及び各方向制御弁のそれぞれの圧力補償をする複数の圧力補償弁とを有し、各前記圧力補償弁は閉じ方向に圧力補償弁の下流側の圧力（Ｐｚ）及び前記複数のアクチュエータのうちの最高負荷圧力（Ｐｍ）をそれぞれ作用させ、圧力補償弁の開方向に圧力補償弁の上流側の圧力であるポンプ吐出圧力（Ｐｄ）及び方向制御弁の下流側圧力であるアクチュエータ負荷圧力（ＰＬ）をそれぞれ作用させて前記圧力補償をし、該可変ポンプの吐出油を該可変ポンプの押しのけ容積変更手段に連通させるポンプ流量調整弁を有し、該最高負荷圧力（Ｐｍ）を油路を介して該ポンプ流量調整弁を閉じ該可変ポンプの押しのけ容積を増大させる方向に該ポンプ流量調整弁のスプリングの作用力とともに作用させ、ポンプ吐出圧力（Ｐｄ）を別の油路を介して該ポンプ流量調整弁を開き該可変ポンプの押しのけ容積を減少させる方向に作用させるようにされた油圧駆動装置において、
前記圧力補償弁の少なくとも１は、そのアクチュエータ自身の自己の負荷圧力の増加に応じてそのアクチュエータに通じる圧力補償弁の少なくとも１の流量を減少するようにしたことを特徴とする油圧駆動装置を提供することによって上記課題を解決した。
【０００８】
【発明の効果】
かかる請求項２、１０記載の構成により、前記圧力補償弁の少なくとも１は、そのアクチュエータ自身の自己の負荷圧力の増加に応じてそのアクチュエータに通じる流量を減少するように即ち方向制御弁差圧が減少するようにしたので、自己負荷圧力の急激な変動があってもアクチュエータ負荷圧力が減衰し油圧制御系が安定し、アクチュエータの最高負荷圧力や前記可変ポンプ吐出圧力に影響を受けない圧力補償弁の圧力補償特性が得られ、単独操作、複合操作のいずれであっても、低負荷側、高負荷側共ハンチングの生じない安定した操作性が得られるという従来例にない優れた効果を奏するものとなった。さらに圧力補償弁の圧力補償特性は圧力補償弁の内部の部品を変更するだけで、そのアクチュエータ自身の自己の負荷圧力の増大に応じて圧力補償弁の流量を減少させる程度を示すいわゆる右下り特性を容易に設定でき、そのアクチュエータ自身の自己の負荷特性に応じてハンチングを生じないような右下り特性を得ることができる。又圧力補償弁の構造そのものが単純であることから厳しい精度が要求されることがなく、コスト安にしかも高い信頼性が得られるものとなった。
【０００９】
【課題を解決するための手段】
本発明の第２発明によると、請求項１記載の発明のように、可変ポンプと、該可変ポンプの吐出油によって駆動される複数の油圧アクチュエータと、該複数のアクチュエータのそれぞれに流入する圧油を制御可能にされた流量調節機能を有する複数の方向制御弁及び各方向制御弁のそれぞれの圧力補償をする複数の圧力補償弁と、前記可変ポンプ吐出圧力（Ｐｄ）とアクチュエータの最高負荷圧力（Ｐｍ）との差圧に等しい二次圧力（Ｐｃ＝Ｐｄ−Ｐｍ）を発生する差圧制御弁と、該可変ポンプの吐出油を該可変ポンプの押しのけ容積変更手段に連通させるポンプ流量調整弁と、を有し、各前記圧力補償弁は圧力補償弁を閉じる方向に圧力補償弁の下流側の圧力（Ｐｚ）を作用させ、圧力補償弁を開く方向に該差圧制御弁から出力される二次圧力（Ｐｃ）及び方向制御弁の下流側圧力であるアクチュエータ負荷圧力（ＰＬ）をそれぞれ作用させて前記圧力補償をするようにした油圧駆動装置において、
該ポンプ流量調整弁のスプリングの作用力を該ポンプ流量調整弁を閉じ該可変ポンプの押しのけ容積を増大させる方向に作用させ、該二次圧力（Ｐｃ）を油路を介して該可変ポンプのポンプ流量調整弁を開き該可変ポンプの押しのけ容積を減少させる方向に作用させたことを特徴とする油圧駆動装置を提供することによって上記課題を解決した。
【００１０】
【発明の効果】
かかる請求項１記載の発明の構成により、二次圧力（Ｐｃ）を油路を介して可変ポンプのポンプ流量調整弁を開き可変ポンプの押しのけ容積を減少させる方向に作用させるようにしたので、従来技術のように、最高負荷圧力（Ｐｍ）をバルブ装置から長く細いパイロット油路を介してポンプの押しのけ容積変更するポンプ流量調整弁を閉じる方向に作用させ、かつポンプ吐出圧力（Ｐｄ）を別の油路を介し前記ポンプ流量調整弁を開く方向に作用させるものに比べ、本発明の第２発明のものは、低温時に作動油の粘度が高くなり、ポンプからバルブ装置までの管路での圧力損失が過大なものとなった場合においても、油路の二次圧力（Ｐｃ）は、バルブ装置内でのポンプ吐出圧力（Ｐｄ）と最高負荷圧力（Ｐｍ）との差圧に等しい二次圧力（Ｐｃ）を発生させていることから、ポンプ吐出管路の圧力損失の大小に関係なく、バルブ装置内のポンプ吐出油路のポンプ吐出圧力（Ｐｄ）が、最高負荷圧力（Ｐｍ）に対しポンプ流量調整弁のスプリングの作用力に相当する圧力に制御されるため、低温時であっても、従来装置と異なり、流量が大幅に減少することがなく、アクチュエータの速度が遅くなることがない。
【００１１】
さらに好ましくは、請求項３、４、８ないし９に記載の発明のように、請求項２記載の発明において、該圧力補償弁の少なくとも１の、そのアクチュエータ自身の自己の負荷圧力の増加に応じてそのアクチュエータに通じる流量を減少するようにさせ、前記圧力補償弁は対応する各前記方向制御弁の上流側に設けられ、該圧力補償弁は第１の油室の第１の受圧面積に自身の下流側の出口圧力を弁を閉じる方向に作用させ、第２の油室の第２の受圧面積に前記二次圧力を弁を開く方向に作用させ、そして第３の油室の第３の受圧面積に各前記アクチュエータの負荷圧力を弁を開く方向に作用させるようにし、そして前記第２と第３の受圧面積をほぼ同じとしかつ前記第１の受圧面積を前記第３の受圧面積より大きくされている。
【００１２】
請求項５及び 12 記載の発明のように、油圧走行車両の一対のクローラを回転駆動させる２個の走行モータでは、前記複数のアクチュエータのうちの少なくとも２個のアクチュエータがアクチュエータの負荷圧力に拘わらず同期して駆動する必要があるため、前記２個のアクチュエータに通ずる２個の前記圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値をそれぞれ等しくすることが望ましい。これにより、左右の走行モータへの供給流量が変化して回転数に差がでると、流量の多い方の負荷圧力が上昇するが、負荷圧力の増加に応じて圧力補償弁の流量が減少するようにされており、かつ左右一対の走行モータの圧力補償弁の流量特性を互いに等しくしたので、方向制御弁のスプールの加工誤差、圧力補償弁の加工誤差による受圧面積の誤差があるときにおいても、かかる誤差は一方の流量の多い方の負荷圧力を上昇させる。吐出圧力と最高負荷圧力との差圧は一定であるから、負荷圧力の上昇により、流量の多い方の圧力補償弁は、方向制御弁差圧を小さくして流量を減じる方向に働くので、流入量が減少し、流量の多い方の走行モータの回転が落ちる。他方の走行モータにおいては、負荷圧力及び吐出圧力と最高負荷圧力との差圧は変化しないので、流量も変化せず、回転数も変化しないので、良好な走行直進性を確保するものとなった。一方、方向転換する場合には、やはり、流量の多い方の走行モータの負荷圧力が上がり、直進性を保持するように働くが、方向転換時には方向制御弁の開度は左右で大きく異なるので補正しきれずに直線性を保てずに、各方向制御弁の操作ストロークに応じて流量が各走行モータに供給され方向転換できる。本発明では、本来の圧力補償弁の改良の他に特別な付属のバルブを必要としないので、バルブ全体の寸法が大きくならないとともに、コストも安く、同時に、使い勝手も良いという優れた効果を奏する。好ましくは、流量の減少率を大きくすると、直進時に過剰に補正され蛇行しやすくなり、また、方向転換時に直進性を保持しようとして、操作しにくくなったりし、逆に流量の減少率が小さいと、補正ができず直進性が確保できないことになるため、前記圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値が０．９９〜０．９５（９９〜９５％）であることが望ましい。
【００１３】
さらに好ましくは、請求項３記載の発明のように、請求項２記載の油圧駆動装置において、前記圧力補償弁の少なくとも１は対応する各前記方向制御弁の上流側に設けられ、該圧力補償弁は第１の油室の第１の受圧面積に自身の下流側の出口圧力を弁を閉じる方向に作用させ、第２の油室の第２の受圧面積に前記二次圧力を弁を開く方向に作用させ、そして第３の油室の第３の受圧面積に各前記アクチュエータの負荷圧力を弁を開く方向に作用させるようにし、そして前記第２と第３の受圧面積をほぼ同じとしかつ前記第１の受圧面積を前記第３の受圧面積より大きくされている油圧駆動装置において、前記複数の油圧アクチュエータのうちの少なくとも２個のアクチュエータのうちの第１のアクチュエータの負荷圧力が他方の第２のアクチュエータの負荷圧力より極めて大きい負荷特性の場合、高負荷側のアクチュエータに通ずる高負荷側圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値を、低負荷側アクチュエータに通ずる低負荷側圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値より、小さくした。かかる構成により、高負荷側アクチュエータの負荷圧力が急上昇した際には、高負荷特性側アクチュエータの流量が減少しその減少した分の流量が低負荷側アクチュエータに供給するようにしたので、低負荷側アクチュエータの速度低下を防止することができる。好ましくは、前記低負荷側アクチュエータの圧力補償弁の第３の受圧面積を第１の受圧面積で除した値が１〜０．９８、前記高負荷特性側アクチュエータの圧力補償弁の第３の受圧面積を第１の受圧面積で除した値が０．９７〜０．９４であることが望ましい。明細書の実体的内容については変更なし。
【００１６】
【発明の実施の形態】
本発明の第１発明の実施形態である油圧駆動装置の油圧回路図について、図１を参照して説明する。エンジン等の原動機１で駆動される可変容量形油圧ポンプ２（以下ポンプという）の吐出油路２３、３に、複数の圧力補償弁４１、４２（うち２個のみ示す）を並列に接続し、各方向制御弁の圧力補償をする各圧力補償弁の出力油路６に夫々チェック弁４０を介して複数のアクチュエータ１０、２０（うち２個のみ示す）それぞれに流入する圧油を制御可能にされた流量調節機能を有する方向制御弁８、１８（うち２個のみ示す）をそれぞれ接続し、それらの方向制御弁の出力側を夫々アクチュエータ１０、２０に接続し、夫々のアクチュエータ１０、２０からの戻り油を再び夫々の方向制御弁８、１８を介してタンク１２へ戻すようにされている。方向制御弁８、１８のアクチュエータ負荷圧力取出ポート７から負荷圧力取出ライン９を介して取り出した負荷圧力は、シャトル弁１３によってアクチュエータ１０、２０のうちの最高負荷圧力（以下最高負荷圧力という）（Ｐｍ）が選択される。圧力補償弁４１、４２は閉じ方向に各圧力補償弁の制御油室に圧力補償弁の下流側の圧力（Ｐｚ）及び最高負荷圧力（Ｐｍ）をそれぞれ作用させ、また、夫々の圧力補償弁４１、４２の開方向に各圧力補償弁の制御油室に圧力補償弁の上流側の圧力であるポンプ吐出圧力（Ｐｄ）及び方向制御弁の下流側圧力であるアクチュエータ負荷圧力（ＰＬ）をそれぞれ作用させるようにしており、圧力補償弁４１、４２はポンプ２の吐出量がアクチュエータ１０、２０の所定要求量を下回った場合には、アクチュエータに適切な比でポンプ２の吐出量を分配するアンチサーチュレーション機能を有する。
【００１７】
さらに、可変ポンプ２の吐出油を可変ポンプの押しのけ容積変更手段１７に連通させるポンプ流量調整弁４５を有し、最高負荷圧力（Ｐｍ）を油路３５を介してポンプ流量調整弁４５を閉じ可変ポンプ２の押しのけ容積を増大させる方向にポンプ流量調整弁のスプリング４６との作用力と共に作用させ、ポンプ吐出圧力（Ｐｄ）を別の油路２３′を介してポンプ流量調整弁４５を開き可変ポンプ２の押しのけ容積を減少させる方向に作用させ、ポンプ吐出圧力Ｐｄと、最高負荷圧力Ｐｍとスプリング４６とであらかじめ設定された作用力と、をつり合わせることにより、ポンプ吐出圧力Ｐｄの作用力が、最高負荷圧力Ｐｍとスプリング４６との作用力、よりも大きい場合は、可変ポンプ２の押しのけ容積を小さくするように、逆にポンプ吐出圧力Ｐｄの作用力が、最高負荷圧力Ｐｍとスプリング４６との作用力、よりも小さい場合は、ポンプ２の押しのけ容積を大きくするように制御されており、これにより（最高）負荷圧力に応じて可変ポンプ２の吐出量を制御するロードセンシング機能を有する。本発明の第１発明では、図１に示す油圧駆動装置において、各アクチュエータの負荷圧力の増加に応じてそのアクチュエータに通じる圧力補償弁４１、４２の少なくとも１の流量を減少するようにしたものである。
【００１８】
かかる構成により、本発明第１発明では圧力補償弁４１、４２の少なくとも１の閉方向油室の受圧面積を開方向油室のそれより大きくして、アクチュエータ負荷圧力の増加に応じてそのアクチュエータに通じる圧力補償弁の流量を減少するように即ち方向制御弁差圧が減少するようにしたので、アクチュエータの負荷圧力の急激な変動があってもアクチュエータ負荷圧力が減衰し油圧制御系が安定し、アクチュエータの最高負荷圧力や前記可変ポンプ吐出圧力に影響を受けない圧力補償弁の圧力補償特性が得られ、単独操作、複合操作のいずれであっても、低負荷側、高負荷側共ハンチングの生じない安定した操作性が得られるという従来例にない優れた効果を奏するものとなった。
【００１９】
本発明の第２発明の実施形態である油圧駆動装置の油圧回路図について、図２（ａ）を参照して説明する。図１の実施形態と同様な部分については同じ符号を付し説明の一部を省略する。図２（ａ）の油圧回路図では、シャトル弁１３がアクチュエータ１０、２０のうちの最高負荷圧力（Ｐｍ）を選択し、可変ポンプ吐出圧力（Ｐｄ）と最高負荷圧力（Ｐｍ）との差圧に等しい二次圧力（Ｐｃ）を発生する差圧制御弁３１がバルブ装置２２内に設けられ、各圧力補償弁４、１４は、圧力補償弁を閉じる方向に圧力補償弁の下流側６の出口圧力（Ｐｚ）を作用させ、圧力補償弁４、１４を開く方向に差圧制御弁３１から取り出した二次圧力ライン３２の二次圧力（Ｐｃ）及び方向制御弁の下流側圧力であるアクチュエータ１０、２０から取り出した負荷圧力ライン３４の負荷圧力（ＰＬ）をそれぞれ作用させるようしている。さらに、可変ポンプ２の吐出油を可変ポンプの押しのけ容積変更手段１７に連通させるポンプ流量調整弁３８を有し、ポンプ流量調整弁のスプリング１９の作用力をポンプ流量調整弁を閉じポンプ２の押しのけ容積を増大させるように作用させ、二次圧力Ｐｃを油路３３を介してポンプ流量調整弁３８を開きポンプ２の押しのけ容積が減少するように作用させ、二次圧力Ｐｃとスプリング１９であらかじめ設定された作用力とをつり合わせることにより、二次圧力Ｐｃが作用する力がスプリング１９の作用力よりも大きい場合は、可変ポンプ２の押しのけ容積を小さくするように、二次圧力Ｐｃが、スプリング１９の作用力よりも小さい場合は、ポンプ２の押しのけ容積を大きくするように容積変更手段１７を作動させるロードセンシング機能を有する。
【００２０】
ここで、図２（ａ）の油圧駆動装置の作用を説明すると、夫々の圧力補償弁４、１４で方向制御弁８、１８の上流側６の圧力が、下流側の夫々のアクチュエータの負荷圧力（ＰＬ）と二次圧力（Ｐｃ）との和とつり合うように夫々の圧力補償弁４、１４が作用することから、夫々の方向制御弁差圧は、仮にそれぞれの受圧面積が等しいとすると、アクチュエータの負荷圧に依らず、前記の二次圧力（Ｐｃ）と等しくなる。すなわち、ポンプ吐出圧力（Ｐｄ）とアクチュエータのうちの最高負荷圧力（Ｐｍ）との差圧に等しくなる。さらに、二次圧力（Ｐｃ）は油路３３でポンプ流量調整弁３８へ導びかれ、ポンプ流量調整弁３８のスプリング１９の作用力とつり合っていることから、ポンプ２の吐出圧力（Ｐｄ）は、二次圧力（Ｐｃ）がスプリング１９の作用力に相当する圧力と等しくなるように制御される。このことは、可変ポンプ吐出圧力（Ｐｄ）は、二次圧力（Ｐｃ）とスプリング１９の作用力に相当する圧力分とが等しくなるように制御される。従って、夫々の方向制御弁８、１８の方向制御弁差圧も、スプリング１９の作用力に相当する圧力に制御される。このような構成にすることにより、いま、仮りにポンプ吐出量が不足した場合には、ポンプ吐出圧力（Ｐｄ）とアクチュエータのうちの最高負荷圧力（Ｐｍ）との差圧に等しい二次圧力（Ｐｃ）は、前述のスプリング１９で設定された差圧分を確保することができなくなるため、夫々の方向制御弁差圧も、設定値よりも小さくなるが、方向制御弁差圧は等しくなるので、夫々のアクチュエータ１０、２０への流量は、方向制御弁８、１８の絞り開度の比率に等しい流量に分流されることになり、アンチサチュレーション機能を有することになる。
【００２１】
かかる構成により、本発明の第２発明では、二次圧力（Ｐｃ）をパイロット油路３３を介して可変ポンプ２のポンプ流量調整弁３８を閉じる方向に、そして可変ポンプ２の押しのけ容積を減少させる方向に作用させるようにしたので、従来技術又は図１のように、最高負荷圧力Ｐｍをバルブ装置４３から長く細いパイロット油路３５を介してポンプ流量調整弁４５を閉じる方向に作用させ、かつポンプ吐出圧力Ｐｄをポンプ流量調整弁４５を開く方向に作用させるものに比べ、本発明の第２発明のものは、低温時に作動油の粘度が高くなり、ポンプ２からバルブ装置２２までの管路２３での圧力損失が過大なものとなった場合においても、油路３２の二次圧力（Ｐｃ）は、バルブ装置２２内でのポンプ吐出路３と最高負荷圧力（Ｐｍ）との差圧（Ｐｃ）に対応した圧力を発生させていることから、ポンプ２からバルブ装置２２までのポンプ吐出管路２３の圧力損失の大小に関係なく、バルブ装置２２内のポンプ吐出油路３のポンプ吐出圧力（Ｐｄ）が、アクチュエータの最高負荷圧力に対しスプリング１９の作用力に相当する圧力に制御されるため、低温時であっても、従来装置と異なり、流量が大幅に減少することがなく、アクチュエータの速度が遅くなることがない。
【００２２】
図２（ａ）の実施例油圧回路においても、本発明の第１発明で開示するように、圧力補償弁４、１４の閉方向油室の受圧面積を開方向油室のそれより大きくして、アクチュエータ負荷圧力の増加に応じてそのアクチュエータに通じる圧力補償弁の流量を減少するようにできる。これにより、自己負荷圧力の急激な変動があってもアクチュエータ負荷圧力が減衰し油圧制御系が安定し、アクチュエータの最高負荷圧力や前記可変ポンプ吐出圧力に影響を受けない圧力補償弁の圧力補償特性が得られ、単独操作、複合操作のいずれであっても、低負荷側、高負荷側共ハンチングの生じない安定した操作性が得られるという従来例にない優れた効果を奏するものとなった。
【００２３】
図３を参照すると、図２の油圧駆動装置に使用した圧力補償弁４、１４の実施形態の断面構造図が示されている。圧力補償弁の断面構造自身は４、１４とも同じであるので、ここでは代表して圧力補償弁４で示す。しかしながら後述のように、圧力補償弁４と１４とは、各圧力室の受圧面積を異なるものとすることもできる。圧力補償弁４は、本体１０１と、本体１０１に設けた小径本体穴１１１とそれに続く大径本体穴１３０の２つの内径を有する本体穴１２８と、小径本体穴１１１（内径ｄ３）に摺動可能に嵌合する小径部１３２及び大径本体穴１３０（内径ｄ２）と摺動可能に嵌合する第１及び第２の大径ランド１３３，１３４を有するスプール１１２と、本体穴１２８に沿って本体１０１に順次設けられたアクチュエータの負荷圧力ポート１０３、二次圧力ポート１０４、出口ポート１０５、ポンプ吐出油路と連通する入口ポート１０２、及びタンクポート１０６とを有する。小径本体穴１１１に嵌合するスプール１１２の一端に設けた小径部１３２はスプリング１１８を介して本体穴端面１２７に当接可能に負荷圧力ポート１０３に通じる第３の油室１１９を形成しそしてスプール１１２の他端１１４はタンクポート１０６に通じる油室１２４を形成する。
【００２４】
スプール１１２の小径部１３２と第１の大径ランド１３３との接合部を囲む大径本体穴１３０内に二次圧力ポート１０４に通じる第２の油室１１３を形成し、スプール１１２の他端１１４に設けられた軸方向穴１１６（内径ｄ１）にピストン１１７が油密に入れ子式に摺動可能に挿入されかつピストン１１７の他端はもう一方の本体穴端面１２６に当接可能にされてタンクポート１０６に通じるタンク油室１２４内にある。軸方向穴１１６内のスプール１１２とピストン１１７との間にパイロット油路１２３を介して出口ポート１０５に通じる第１の油室１２１を形成しており、第１の油室１２１の第１の受圧面積Ａ１はピストン１１７の断面積により、第２の油室１１３の第２の受圧面積Ａ２は大径本体穴１３０断面積から小径穴１１１断面積を引いた面積により、そして第３の油室１１９の第３の受圧面積Ａ３は小径部１３２断面積により、それぞれ形成させ、かつスプール１１２には、第１の大径ランド部１３３に面する第２の大径ランド１３４に設けた出口ポート１０５と入口ポート１０２間を絞る開閉可能な絞り部１１５とを有する。出口ポート１０５に通じる第１の油室１２１には出口圧力Ｐｚがスプール１１２を図で見て左方向に絞り部１１５を閉じる方向に作用し、第２の油室１１３の第２の受圧面積Ａ２には二次圧力Ｐｃがスプール１１２を図で見て右方向に絞り部１１５を開く方向に作用し、そして第３の油室１１９の第３の受圧面積Ａ３には負荷圧力ＰＬがスプール１１２を図で見て右方向に絞り部１１５を開く方向に作用する。
【００２５】
図３の実施態様では、第３の受圧面積Ａ３と第２の受圧面積Ａ２がほぼ等しくなるようにされ、スプール１１２小径部１３２の外径ｄ３をピストン１１７の外径ｄ１に対しやや小さく（ｄ３＜ｄ１）して第３の受圧面積Ａ３が第１の受圧面積Ａ１より小さくされている。また、スプール１１２は、図３で見て、左方向へ最大ストロークした場合は、スプールの左端面が第３の油室１１９の本体左端面１２７に当接し、絞り部１１５を閉じるようにされている。逆に右方向へ最大ストロークした場合は、スプールの右端面１１４及びピストン１１７の右端面が本体穴右端面１２６に当接し、絞り部１１５は全開となるようにされている。スプール１１２の中間のストロークでは、スプールの絞り部１１５によりスプールの右方向へのストローク量に比例して、開度が比例的に増加するようにされている。なお、スプリング１１８は、方向制御弁８、１８が操作されていない時にスプール１１２を右方向へストロークさせ、絞り部１１５を開いておくためのものであり、その作用力は極く弱いものである。また図３は作動原理を概念的に示すためのものであり、本体穴１２８の両端は開放されていないが、実際には本体穴を図示しない段付きの通し穴もしくは右側面からの加工穴として構成し、図示しないねじプラグ等の方法で閉止する構造とすることができる。
【００２６】
次に図３の実施形態についてその作用を説明する。圧力補償弁のスプール１１２に作用する力のバランスを考える。まず、スプール１１２を同図右方向すなわち、絞り部１１５を開く方向に作用する力は、負荷圧力をＰＬ、ポンプ吐出圧力をＰｄ、最高負荷圧力をＰｍ、二次圧力をＰｃ（Ｐｃ＝Ｐｄ−Ｐｍ）とすると、
（Ａ３・ＰＬ）＋（Ａ２・Ｐｃ）………（１）
逆に同図左方向、すなわち絞り部１１５を閉じ方向に作用する力は、方向制御弁上流側６、即ち出口ポート１０５の出口圧力をＰｚとすると、
（Ａ１・Ｐｚ）………（２）
となる。ここで圧力補償弁の制御時は両方向の力がつり合っているので（１）式と（２）式は等しいから
（Ａ３・ＰＬ）＋（Ａ２・Ｐｃ）＝（Ａ１・Ｐｚ）………（３）
なる関係が成立する。但しスプリング１１８の作用力はごく弱いため無視する。
仮にここでスプール小径部外径ｄ３とピストン１１７の外径ｄ１が等しいとすると、Ａ３＝Ａ１となり、（３）式より方向制御弁差圧ΔＰ＝（Ｐｚ−ＰＬ）は、
ΔＰ＝（Ｐｚ−ＰＬ）＝（Ａ２／Ａ３）・Ｐｃ………（４）
となる。従って方向制御弁差圧ΔＰは、二次圧力Ｐｃと、スプール１１２の外径ｄ２と外径ｄ３とピストン１１７の外径ｄ１の寸法により一定値に決定されるから、個々の負荷圧力ＰＬによらず常に一定の値となる。なお、サチュレーション状態では、二次圧力Ｐｃがその状況に応じて小さくなり、方向制御弁差圧も小さくなるが、前述したように、差圧はそれぞれ等しいので、夫々のアクチュエータ１０、２０への流量は、方向制御弁８、１８の絞り開度の比率に等しい流量に分流されることになり、アンチサチュレーション機能を有することとなり、夫々のアクチュエータ１０、２０への流量は、個々の負荷圧力ＰＬの影響は受けない。なお、第３の受圧面積Ａ３と第２の受圧面積Ａ２は等しくても等しくなくてもよい。仮に、Ａ２＝Ａ３の場合はΔＰ＝Ｐｃであり、Ａ２≠Ａ３の場合は（４）式に示すようにＡ２とＡ３の比率によってΔＰの絶対値を種々変更できる。なお、第１の受圧面積は第３の受圧面積との関係で決定される。
【００２７】
さて、本発明の第１発明においては、ピストン１１７の外径ｄ１をスプール１１２の小径部外径ｄ３に対しやや大きく（ｄ１＞ｄ３）している。従って、Ａ３＝ｋ・Ａ１（但しｋ＜１）として（３）式に代入すると
ｋ・Ａ１・ＰＬ＋Ａ２・Ｐｃ＝Ａ１・Ｐｚ………（５）
（５）式において便宜上、ｋ＝｛１−（１−ｋ）｝とおくと、
｛１−（１−ｋ）｝・Ａ１・ＰＬ＋Ａ２・Ｐｃ＝Ａ１・Ｐｚ
となり、変形して、
Ａ１・ＰＬ−Ａ１・（１−ｋ）・ＰＬ＋Ａ２・Ｐｃ＝Ａ１・Ｐｚ
ＰＬ−（１−ｋ）・ＰＬ＋（Ａ２／Ａ１）・Ｐｃ＝Ｐｚ
−（１−ｋ）・ＰＬ＋（Ａ２／Ａ１）・Ｐｃ＝Ｐｚ−ＰＬ
よって方向制御弁差圧ΔＰは
ΔＰ＝（Ｐｚ−ＰＬ）＝（Ａ２／Ａ１）・Ｐｃ−（１−ｋ）・ＰＬ………（６）
または、Ａ１＝Ａ３／ｋを代入すると。
ΔＰ＝〔（ｋ・Ａ２）／Ａ３〕・Ｐｃ−（１−ｋ）・ＰＬ………（７）
を得る。ここでは、定数ｋは１より小さい値であるから、（６）、（７）式の右辺の第２項は負の値となる。（６）、（７）式によれば、方向制御弁差圧ΔＰは二次圧力Ｐｃ及びアクチュエータ負荷圧力ＰＬの一次式となり、かつアクチュエータ負荷圧力ＰＬの増大に応じて、夫々の方向制御弁差圧ΔＰが減少し流量が減少する。即ち、アクチュエータ負荷圧力ＰＬの増大に応じて流量が減少する右下りの圧力補償特性が得られる。
これは一の方向制御弁のみを操作した場合であっても、二以上の方向制御弁を同時に操作した場合であっても、ポンプの最大吐出量が全てのアクチュエータの必要流量以上の条件、即ちサチュレーション状態に至っていない場合であれば常に成立する。即ち、この状態では、前述したように二次圧力Ｐｃはスプリング１９の作用力で設定した一定の圧力に保たれることになる。それに対して負荷圧力ＰＬは各々のアクチュエータ負荷圧力であるため、他のアクチュエータ負荷圧力やアクチュエータの最高負荷圧力又はポンプ吐出圧力によらず一貫してそれぞれのアクチュエータ負荷圧力のみに依存した右下がりの圧力補償特性が得られる。一方、ポンプ吐出量が不足したサチュレーション状態においては、二次圧力Ｐｃはスプリング１９で設定した作用力より小さな圧力Ｐｃ′となり、そのＰｃ′の大きさは流量の不足度合いにより左右され一定の値にならない。しかし、全ての圧力補償弁には等しい二次圧力Ｐｃ′が作用しており、ポンプの吐出量の全てが各々のアクチュエータに適切な比率で分配されることになる。
また、仮に一の方向制御弁のみが操作されサチュレーション状態になった場合は当然のことながらアクチュエータ負荷圧力ＰＬによらず全ての吐出量が一のアクチュエータに流れることになる。
【００２８】
次に二個の方向制御弁を操作しサチュレーション状態に至った場合について述べる。説明を簡単にするため双方の方向制御弁の開度には変化がないものとし、この状態で一方の側のアクチュエータの負荷圧力のみが上昇し他方の側の負荷圧力には変化がないものとする。負荷圧力が上昇した側の方向制御弁差圧ΔＰは（６）、（７）式によりアクチュエータ負荷圧力ＰＬの上昇に伴い減少する。但し第１項の差圧は小さな圧力Ｐｃ′であるため流量そのものは小さい。負荷圧力が変化しない他方の側の方向制御弁差圧はアクチュエータ負荷圧力ＰＬが変化しないため（６）（７）式の第２項には変化がないが、負荷圧力が上昇した側の流量減少により全体的に流量に余裕が生じ二次圧Ｐｃ′が上昇するため第１項が大きくなり、方向制御弁差圧が大きくなり流量が増加する。言い換えれば、全体としてポンプ吐出量の全てはアクチュエータに分配され、負荷圧力が上昇した側の流量が減少した分だけ負荷圧力が変化しない側の流量が増加する。従ってサチュレーション状態では負荷圧力が変化しない側の流量が負荷圧力が上昇する側の負荷圧力により自己の負荷圧力が一定であるにも係わらず増加することになるが、元来サチュレーション状態では要求される流量に対して実際に流れる流量が不足しているため目標速度に対するオーバシュートが発生しないためこの特性はハンチングの原因とはなりえない。むしろ負荷圧力が上昇しない側の流量が増加するのは不足している流量を補いより目標に近い速度にできるという利点が生じる。
【００２９】
一方、他方の負荷圧力が減少した場合は上記とは逆の現象が起きる。すなわち、負荷圧力の減少した側は流量が増加し負荷圧力が一定の側は流量が減少するということになる。また、負荷圧力の変動が同じ割合で上昇又は下降した場合はそれぞれの流量の分流比は変化のないまま推移することになる。これは、三以上の方向制御弁を同時に操作した場合にも成立する。このように、ポンプ吐出量が充分な場合であっても、サチュレーション状態であっても本発明によればハンチングの生じない常に安定した制御性が得られる。
さらに、当然のことながら、定数ｋの値を変えることによって圧力補償特性を任意の値に設定することが可能となる。すなわち、ｋの値を小さくすればする程、圧力補償弁の右下りの度合いが大きくなる。これは、各々のアクチュエータの負荷特性に応じて、右下りの度合を設定できることを意味する。しかも、その設定は、具体的にはピストン１１７の外径ｄ１を変えるだけで済み、本体１０１そのものを変える必要はなく、簡単に変更可能である。
なお、定数ｋの値は、実際の装置に合わせて決定されるが、ハンチングを起きやすいアクチュエータにおいては、補償流量の減少率が小さいとハンチングしやすくなり、減少率が大きいと本来の流量を一定とする圧力補償機能を得られないので、０．９９＞ｋ＞０．９５（０．９９〜９５％）程度が適当である。このように、右下がりの度合いばかりでなく、同一の本体で種々のｋの値を簡単に得られるので、負荷条件に応じた種々の圧力補償弁が簡単に得られる。
【００３０】
図７（ａ）に示すように、図２（ａ）に示す回路を有する油圧駆動装置において、油圧走行車両の一対のクローラを回転駆動させる２個の走行モータ１５、１５′駆動のように、複数のアクチュエータのうちの少なくとも２個のアクチュエータ１５、１５′がアクチュエータの負荷圧力に拘わらず同期して駆動する必要のあるときに、前記２個のアクチュエータ１５、１５′に通ずる２個の圧力補償弁２８、２９の第３の受圧面積Ａ３を第１の受圧面積Ａ１で除した値をそれぞれ等しくすることが望ましい。かかる構成により、左右の走行モータ１５、１５′の供給流量が変化して回転数に差がでると、供給流量の多い方の負荷圧力が上昇するが、負荷圧力の増加に応じて圧力補償弁の流量が減少するようにされており、かつ左右一対の走行モータの圧力補償弁の流量特性を互いに等しくしたので、方向制御弁８、１８のスプールの加工誤差、圧力補償弁の加工誤差による受圧面積の誤差があるときにおいても、かかる誤差は一方の流量の多い方の負荷圧力ＰＬを上昇させる。ここで吐出圧力Ｐｄと最高負荷圧力Ｐｍとの差圧Ｐｃは一定であるから、負荷圧力の上昇により、流量の多い方の圧力補償弁は、方向制御弁差圧を小さくして流量を減じる方向に働くので、流入量が減少し、流量の多い方の走行モータの回転が落ちる。他方の走行モータにおいては、負荷圧力及び吐出圧力と最高負荷圧力との差圧は変化しないので、流量も変化せず、回転数も変化しないので、良好な走行直進性を確保するものとなった。一方、方向転換する場合には、やはり、流量の多い方の走行モータの負荷圧力が上がり、直進性を保持するように働くが、方向転換時には方向制御弁の開度は左右で大きく異なるので補正しきれずに直線性を保てずに、各方向制御弁の操作ストロークに応じて流量が各走行モータに供給され方向転換できる。本発明では、本来の圧力補償弁の改良の他に特別な付属のバルブを必要としないので、バルブ全体の寸法が大きくならないとともに、コストも安く、同時に、使い勝手も良いという優れた効果を奏する。
【００３１】
好ましくは、流量の減少率を大きくすると、直進時に過剰に補正され蛇行しやすくなり、また、方向転換時に直進性を保持しようとして、操作しにくくなったりし、逆に流量の減少率が小さいと、補正ができず直進性が確保できないことになるため、前記圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値が０．９９〜０．９５（９９〜９５％）であることが望ましい。
【００３２】
図７（ｂ）に示すように、さらに好ましくは、本発明の第２発明の油圧駆動装置において、複数の油圧アクチュエータのうちの少なくとも２個のアクチュエータ１１、２５のうちの旋回用油圧モータのような高負荷側アクチュエータ２５の負荷圧力が他方のブーム用油圧シリンダのような低負荷側アクチュエータ１１の負荷圧力より極めて大きい負荷特性の場合、好ましくは、高負荷側のアクチュエータ２５に通ずる高負荷側圧力補償弁３６の第３の受圧面積Ａ３を前記第１の受圧面積Ａ１で除した値を、低負荷側アクチュエータ１１に通ずる低負荷側圧力補償弁３０の前記第３の受圧面積Ａ３を前記第１の受圧面積Ａ１で除した値より、小さくした。かかる構成により、高負荷特性側アクチュエータ２５の負荷圧力が急上昇した際には、高負荷特性側アクチュエータの流量が減少しその減少した分の流量が低負荷側アクチュエータ１１に供給するようになるので、低負荷側アクチュエータ１１の速度低下を防止することができる。なお好ましくは、低負荷側アクチュエータ１１の圧力補償弁３０の第３の受圧面積Ａ３を第１の受圧面積Ａ１で除した値が１〜０．９８とし、高負荷側アクチュエータ２５の圧力補償弁３６の第３の受圧面積Ａ３を第１の受圧面積Ａ１で除した値が０．９７〜０．９４であることが望ましい。
明細書の実体的内容については変更なし。
【００３３】
さらに、高負荷側アクチュエータ２５の旋回負荷圧力が過大な場合は、圧力補償弁３６の開度は絞られ旋回モータ２５側に供給される流量が減少するので、旋回モータ２５側から図示しないオーバーロードリリーフバルブから流失する無駄なリリーフ流量が少なくなると同時に、旋回モータ２５の負荷圧力そのものの上昇も抑制する。従って、無駄なリリーフ流量が少なくなる分だけ低負荷側アクチュエータ１１であるブーム用油圧シリンダの速度の低下は防止できる。その後旋回モータ２５の速度が増加し加速度が減少するに従い負荷圧力も減少するので徐々に圧力補償弁３６の開度が大きくなり、流量は旋回負荷圧力の減少に伴い徐々に増加していくことになるので緩やかな旋回モータ２５の加速が得られる。さらに、旋回モータの加速が終了し、定常速度による旋回となると、旋回負荷圧力が急激に低下し、ブームシリンダの負荷圧力の方が大きくなる。この時、旋回モータ２５側の圧力補償弁３６の開度は全開ではなく絞りの状態から徐々に開度を増していく途中の過程であって、依然として絞られた状態にある。このため、旋回モータ２５の負荷圧力が急激に低下したとしても、旋回モータ２５側の圧力補償弁３６の開度が急激に小さくなる割合は減少し、旋回モータ２５がショックを伴って減速するということがない。逆に、ブームシリンダ１１側の圧力補償弁３０の開度は、旋回モータ２５の旋回初期段階から二次圧力がある程度確保されているので、比較的大きな開度を確保できていることにより、その後旋回モータの加速が終了し、定常速度による旋回となっても、従来例のように急激に開度が増加することがなく、ショックを伴って加速するということがなくなった。
【００３４】
図４を参照して、図２の油圧駆動装置に使用した図３とは異なる実施形態の圧力補償弁４′について説明する。図３の実施形態と同様な部分については同じ符号を付し説明の一部を省略する。図４の圧力補償弁４′は図３の実施形態に対し、圧力補償弁の開き方向に作用する受圧面積Ａ３、Ａ２の構成を変えたものである。即ち図４では、本体２０１側は本体穴２２８は大径穴（内径ｄ２）だけとし、本体穴２２８には第１、第２及び第３の大径ランド２０９、２１０、２１１を有するスプール２１２が嵌入されている。そして図３の小径穴１１１（内径ｄ３）の替わりに外径をｄ３とした補助ピストン２１７をスプール２１２の外方端２１４に設けた補助軸方向穴２０２に入れ子式に摺動自在に内挿したものである。本体穴２２８に沿って本体２０１に順次設けられた二次圧力ポート２０４、アクチュエータの負荷圧力ポート２０３、出口ポート１０５、ポンプ吐出油路と連通する入口ポート１０２、及びタンクポート１０６とを有する。補助ピストン２１７の外方端は本体穴２２８の端面２２７に当接可能にされて二次圧力ポート２０４に通じる第２の油室２１３を形成する。補助軸方向穴２０２内のスプール２１２と補助ピストン２１７との間にスプリング２１８が介されかつ補助パイロット油路２２３を介して負荷圧力ポート２０３に通じる第３の油室２２０を形成する。そして第１の油室１２１の第１の受圧面積Ａ１はピストン１１７の断面積により、第２の油室２１３の第２の受圧面積Ａ２は本体穴２２８の断面積から補助ピストン２１７の断面積を引いた面積により、そして第３の油室２２０の第３の受圧面積Ａ３は補助ピストン２１７の断面積により、それぞれ形成させる。
【００３５】
かかる構成によれば、各々の径ｄ１，ｄ２，ｄ３の関係は図３の実施形態と同様にしておけば、二次圧力Ｐｃ、即ちポンプ吐出圧力とアクチュエータの最高負荷圧力との差圧はポンプ流量調整弁３８のスプリング１９の作用力となるので、二次圧力Ｐｃに対して、負荷圧力ＰＬが充分大きいので、補助ピストンは本体穴左端面に押しつけられ図３で示した実施形態と同様の作用が得られる。
【００３６】
なお、図４の実施形態においては、自走負荷のような負荷がマイナスになるなどして二次圧力Ｐｃに対して、負荷圧力ＰＬが小さくなった場合は、補助ピストン２１７が本体穴左端面２２７より離れてスプール２１２を付勢するように働き、負荷圧力ＰＬが作用する受圧面積Ａ３に、二次圧力Ｐｃが作用することとなる。この場合は負荷圧力がＰｃと同一とみなされて制御されることになり、方向制御弁の差圧が若干大きくなり流量がやや多目になる。しかし、段付き穴を有する図３の実施形態に対して、図４の実施形態においては、穴加工を段付きにする必要がなく加工が容易となる利点があり、また、二次圧力Ｐｃはもともと小さいため二次圧力Ｐｃが補助ピストン外方端面に作用することに起因する実使用上の問題が少なく、さらに、負荷圧力そのものが常にある一定以上の圧力である装置にはより適したものとなる。
【００３７】
以上、図３及び図４に示す圧力補償弁４、４′は図２及び図７に使用されるものとして説明したが、本発明の油圧駆動装置に使用される圧力補償弁４、４′は図２及び図７の回路例以外にも、適用可能である。すなわち、前述のように開き方向に負荷圧力ＰＬと二次圧力Ｐｃが作用し、閉じ方向に方向制御弁の上流側圧力（圧力補償弁の下流側）Ｐｚが作用し、制御している形態の圧力補償弁であればよい。例えば図２（ｂ）に示す回路に使用できる。
【００３８】
そこでは図２（ａ）に示すポンプ制御回路に二次圧Ｐｃをパイロット油路３３を介して作用させる代わりに、図２（ｂ）では（図１で示すような）可変ポンプ２の吐出油を可変ポンプの押しのけ容積変更手段１７に連通させるポンプ流量調整弁４５を有し、最高負荷圧力（Ｐｍ）を油路３５を介してポンプ流量調整弁４５を閉じ可変ポンプ２の押しのけ容積を増大させる方向に作用させ、ポンプ吐出圧力（Ｐｄ）を別の油路２３′を介してポンプ流量調整弁４５を開き可変ポンプ２の押しのけ容積を減少させる方向に作用させ、ポンプ吐出圧力Ｐｄの作用力を、最高負荷圧力Ｐｍとスプリング４６とであらかじめ設定された作用力と、をつり合わせることにしてもよい。
【００３９】
図２（ｃ）に示すように、図２（ａ）において、二次圧力Ｐｃを生成するものとして、ポンプ吐出圧力取出ライン２３′からの可変ポンプ吐出圧力（Ｐｄ）と最高負荷圧力取出ライン１６からのアクチュエータの最高負荷圧力（Ｐｍ）との差圧を差圧検出器６０で検出し、差圧検出器６０の出力を制御装置６１に入力して制御信号６２を発生させかつ出力させ、制御信号６２で作動する電磁比例弁６３が出力する二次圧力（Ｐｃ）によって、アンチサチュレーション機能を確保してもよい。６４はパイロットポンプである。
【００４０】
図６は図２（ａ）とは異なる油圧駆動装置の実施形態を示す油圧回路図で、方向制御弁の下流側に圧力補償弁を配置した点は、特開平４−１９４０９号公報と同様である。図２（ａ）と同様の部分については同じ符号を付し説明の一部を省略する。図６に示すように、５０、５１は油圧アクチュエータで、吐出油路３の圧油をチェック弁４０、流量制御機能を有する方向制御弁５３、５４、及び圧力補償弁４４、４８を介して導き、アクチュエータ５０、５１の戻り油は方向制御弁５３、５４からタンク油路１２を介してタンクＴに戻される。アクチュエータの負荷圧力のうち最高負荷圧力をシャトル弁１３で選択して双方の圧力補償弁４４、４８を閉じ方向にスプリング４４ａ，４８ａとともに作用させ、方向制御弁５３、５４の下流側圧力Ｐｄ′を圧力補償弁を開く方向にそれぞれ作用させている。方向制御弁５３、５４前後の差圧は図２（ａ）と同様に可変ポンプ吐出圧力Ｐｄと最高負荷圧力Ｐｍとの差圧に一致させるべく構成することにより、アンチサチュレーション機能をもたせるようにしている。バルブ装置２４内の吐出油路３のポンプ吐出圧力Ｐｄとシャトル弁１３で選択された出口油路１６の最高負荷圧力Ｐｍとの差圧に対応した圧力Ｐｃを発生する差圧制御弁３１を図２（ａ）と同様にバルブ装置２４内に設け、差圧制御弁３１で発生した二次圧力Ｐｃを二次圧力ライン３２、パイロット油路３３を介して、ポンプ装置２１のポンプ流量調整弁３８に可変ポンプ２の吐出量を開き押しのけ容積変更手段１７を作動させ可変ポンプ２の吐出量を減少するように作用させ、ポンプ流量調整弁３８のスプリング１９をポンプ流量調整弁を閉じ可変ポンプ２の吐出量を増大するように制御している。
【００４１】
図６の油圧駆動装置の作用を説明すると、可変ポンプ２の吐出口の圧力は、バルブ装置２４内の吐出導管３の圧力に対して吐出導管２３での圧力損失に相当する分だけ高い圧力となることから、吐出油路３のポンプ吐出圧力Ｐｄは最高負荷圧力Ｐｍとスプリング１９の作用力によってのみ決まることになり、作動油の温度に依存しない。すなわち、ポンプ流量調整弁３８内の力のバランス式は、
Ｐｃ＝スプリング１９の作用力
であり、２次圧力Ｐｃは差圧制御弁３１より、
Ｐｃ＝Ｐｄ−Ｐｍであることから、
Ｐｄ−Ｐｍ＝スプリング１９の作用力
が成立し、ポンプ吐出圧力は、
Ｐｄ＝Ｐｍ＋スプリング１９の作用力
となる。
また、圧力補償弁４４、４８内の力のバランスから、
Ｐｄ′＝Ｐｍ＋スプリング４４ａの作用力
となる。
従って、方向制御弁５３、５４前後の差圧＝Ｐｄ−Ｐｄ′＝スプリング１９の作用カースプリング４４ａの作用力
となり、方向制御弁前後の差圧は、ポンプ流量調整弁３８のスプリング１９の作用力と圧力補償弁４４、４８のスプリング４４ａ、４８ａの作用力によってのみ決まり、アクチュエータ５０、５１の負荷圧力には依存しないので、図２（ａ）と同様な、作動油の温度に左右されない油圧駆動装置とすることができる。
一方、サチュレーション状態の場合は、ポンプ吐出量が不足するため上記式の関係は成立しない。しかし、方向制御弁５３、５４の下流側の圧力Ｐｄ′は、最高負荷圧力Ｐｍと圧力補償弁のスプリング４４ａ、４８ａのスプリングの作用力の和となり、各スプリングが同じ場合には、全ての方向制御弁の下流側の圧力は同一となる。一方、全ての方向制御弁の上流側は吐出導管３に並列に通じているため、その圧力はＰｄと同一である。従って、全ての方向制御弁前後の差圧はそれぞれ同一となり、図２（ａ）の場合と同様に可変ポンプ吐出量は夫々の方向制御弁の開度の比率に比例した割合で分流され、アンチサチュレーション機能を有することとなる。
【００４２】
図５を参照して、本発明の第１発明の実施例である図１に使用される圧力補償弁４１について説明する。圧力補償弁４１のボデイ３０１は、第１ボデイ３０１ａと第２ボデイ３０１ｂに分割され、適宜ボルト締め等の方法で（図示せず）一体に組付けてなり、第１ボデイ３０１ａには小径穴３２１と小径穴に続く中径穴３２２とが設けられ、小径穴３２１に嵌合する第１スプール３１１と中径穴３２２に嵌合する第２スプール３１２とが配置され、第２ボデイ３０１ｂには中径穴３２２に続く大径穴３２３と大径穴に続く小径穴３２１と同径の補助小径穴３２５とが設けられ、大径穴３２３及び補助小径穴３２５に嵌合する第３スプール３１０は大径穴３２３に嵌合する第１及び第２の大径ランド３１３，３１４と補助小径穴３２５に嵌合する補助小径部３１５とを有し、第１スプール３１１と小径穴３２１端面３２０間に前記各スプールを押すスプリング３５０を介している。さらに、ボデイ３０１に沿って、ポンプ吐出油路３と連通し小径穴３２１に通じる補助入口ポート３４１、アクチュエータの負荷圧力ライン３４に連通しかつ中径穴３２２に通ずるアクチュエータの負荷圧力ポート３４２、第２スプール３１２と第３スプール３１０との当接部を囲む大径穴３２３に通じるタンクポート３４３、第１及び第２の大径ランド３１３，３１４間の大径穴３２３に通じる出口ポート３４４、ポンプ吐出油路３と連通しかつ第２の大径ランド３１４に設けた開閉可能な絞り部３１６で絞られるよう開口する入口ポート３４５、及びアクチュエータのうちの最高負荷圧力取出ライン１６と連通しかつ第２の大径ランド３１４と補助小径部３１５との連続部の大径穴３２３に通ずる最高負荷圧力ポート３４６、が順次設けられている。補助小径部３１５と補助小径穴端面３３０間にパイロット油路３５１を介して出口ポート３４４と連通する油室３３４を設けている。第１ボデイ３０１ａと第２ボデイ３０１ｂは適宜ボルト締め等の方法で（図示せず）一体に組付けてボデイ３０１を構成するが、この際第１ボデイ３０１ａ側中径穴３２２と第２ボデイ３０１ｂ側大径穴３２３とが芯ずれしていても、第２スプール３１２と第３スプール３１０は別部品で単に当接しているだけであることから、作動上の問題はない。
【００４３】
そして、圧力補償弁４１は閉じ方向に出口ポート３４４の出口圧力（Ｐｚ）をパイロット油路３５１を介して油室３３４の補助小径部３１５の端面３４０（受圧面積Ｂ１）に、最高負荷圧力ポート３４６の最高負荷圧力（Ｐｍ）を第２の大径ランド３１４の断面積から補助小径部３１５の断面積を差し引いた油室３３６の受圧面積Ｂ２に、それぞれ作用させる。さらに圧力補償弁４１を開く方向に補助入口ポート３４１を介してポンプ吐出圧力（Ｐｄ）を油室３３１の第１スプール３１１の受圧面積Ｂ１に、負荷圧力ポート３４２のアクチュエータ負荷圧力（ＰＬ）を第２スプール３１２の断面積から第１スプール３１１の断面積Ｂ１を差し引いた油室３３２の受圧面積Ｂ３に、それぞれ作用させる。なお、第１の大径ランド３１３の断面積から第２スプール３１２の断面積を引いた分の油室３３３の断面積は、タンクポート３４３によりタンクに通じているため、前記各スプールを開閉させる作用力は働らかない。
そして前記断面積Ｂ２と第１スプール３１１の断面積Ｂ１とをほぼ同じとし（Ｂ１＝Ｂ２）、加えて前記断面積Ｂ３は第１スプールの断面積Ｂ１（＝Ｂ２）より小にし（Ｂ１＞Ｂ３）、各アクチュエータの負荷圧力（ＰＬ）の増加に応じてそのアクチュエータに通じる圧力補償弁４１の流量を減少するようにしたものである。
【００４４】
作動において、各スプールを押すスプリング３５０の力は図３のスプリング１１８と同様に弱いものであるのでこの力を無視して、
圧力補償弁４１の各スプールが均衡状態にあるときは、各スプールにかかる力のバランスは、
Ｐｚ・Ｂ１＋Ｐｍ・Ｂ２＝Ｐｄ・Ｂ１＋ＰＬ・Ｂ３………（８）
Ｂ１＝Ｂ２であるから、
Ｐｚ・Ｂ１＋Ｐｍ・Ｂ１＝Ｐｄ・Ｂ１＋ＰＬ・Ｂ３………（９）
Ｐｚ＋Ｐｍ＝Ｐｄ＋ＰＬ・（Ｂ３／Ｂ１）
Ｂ１＞Ｂ３であり、（Ｂ３／Ｂ１）＝ｋとおくと、
Ｐｚ＋Ｐｍ＝Ｐｄ＋ＰＬ・ｋ ………（１０）
ただしｋ＜１ｋ＝１−（１−ｋ）とおくと、
Ｐｚ＋Ｐｍ＝Ｐｄ＋ＰＬ・〔１−（１−ｋ）〕
Ｐｚ＋Ｐｍ＝Ｐｄ＋ＰＬ−ＰＬ・（１−ｋ）………（１１）
ΔＰ＝Ｐｚ−ＰＬであるから、式（１１）より、
Ｐｚ−ＰＬ＝−Ｐｍ＋Ｐｄ−ＰＬ・（１−ｋ）
ΔＰ＝Ｐｚ−ＰＬ＝（Ｐｄ−Ｐｍ）−ＰＬ・（１−ｋ）………（１２）
ここで、図３のＰｃは、Ｐｃ＝Ｐｄ−Ｐｍであるから、
図５の圧力補償弁４１も、図３の圧力補償弁４と同様な作動をする。
【００４５】
図３の圧力補償弁４で説明したと同様な理由で、２個の圧力補償弁４１、４２の第３の受圧面積Ｂ３を第１の受圧面積Ｂ１で除した値が０．９９〜０．９５（９９〜９５％）であることが望ましい。本発明の第１発明の油圧駆動装置の圧力補償弁４１、４２においても、油圧走行車両の一対のクローラを回転駆動させる２個の走行モータ駆動のように、前記複数のアクチュエータのうちの少なくとも２個のアクチュエータがアクチュエータの負荷圧力に拘わらず同期して駆動する必要のあるときは、前記２個のアクチュエータに通ずる２個の圧力補償弁４１、４２の第３の受圧面積Ｂ３を第１の受圧面積Ｂ１で除した値をそれぞれ等しくすることが望ましい。
好ましくは、流量の減少率を大きくすると、直進時に過剰に補正され蛇行しやすくなり、また、方向転換時に直進性を保持しようとして、操作しにくくなったりし、逆に流量の減少率が小さいと、補正ができず直進性が確保できないことになるため、２個の圧力補償弁４１、４２の第３の受圧面積Ｂ３を第１の受圧面積Ｂ１で除した値が０．９９〜０．９５（９９〜９５％）であることが望ましい。図３の圧力補償弁４で説明したと同様な理由で、本発明の第１発明の油圧駆動装置に使用される圧力補償弁４１、４２において、複数のアクチュエータ１０、２０のうちの少なくとも２個のアクチュエータのうちの高負荷側アクチュエータ２０の負荷圧力が他方の低負荷側アクチュエータ１０の負荷圧力より極めて大きいときに、高負荷側アクチュエータ２０に通ずる高負荷側圧力補償弁４２の第３の受圧面積Ｂ３を第１の受圧面積Ｂ１で除した値を、低負荷側アクチュエータ１０に通ずる低負荷側圧力補償弁４１の第３の受圧面積Ｂ３を第１の受圧面積Ｂ１で除した値より、小さくすることが望ましい。そして前記低負荷側アクチュエータの圧力補償弁の第３の受圧面積を第１の受圧面積で除した値が１〜０．９８、前記高負荷側アクチュエータの圧力補償弁の第３の受圧面積を第１の受圧面積で除した値が０．９７〜０．９４であることが望ましい。
【００４６】
以上各実施態様では、２個の油圧アクチュエータを駆動する油圧回路で説明したが、例えば油圧ショベルでは油圧走行車両の一対のクローラを回転駆動させる２個の走行モータ、旋回用油圧モータ、及びブーム、アーム及びバケットの各油圧シリンダーのように、少なくとも６個のアクチュエータが操作される。従って以上各実施態様はこれらを代表した２個のアクチュエータのみが示されており、本発明では、複数の油圧アクチュエータというときは、ぞれぞれ複数の走行モータ、旋回用油圧モータ及び油圧シリンダー等を含み、かつ油圧アクチュエータにそれぞれ通じるぞれぞれ複数の圧力補償弁及び方向制御弁を含むものと理解さるべきである。
【図面の簡単な説明】
【図１】本発明の第１発明の実施形態である油圧駆動装置の油圧回路図。
【図２】（ａ）は本発明の第２発明の実施形態である油圧駆動装置の油圧回路図を示し、（ｂ）は（ａ）とは異なる実施形態であるポンプ吐出量制御回路を示す部分油圧回路図、（ｃ）は（ａ）とは異なる二次圧力発生装置を示す部分油圧回路図で図２（ａ）の他の部分はそのままである。
【図３】図２及び図７の油圧駆動装置に使用される圧力補償弁の実施形態の断面構造図（概念図）を示す。
【図４】図２及び図７の油圧駆動装置に使用される図３とは異なる実施形態の圧力補償弁の断面構造図（概念図）を示す。
【図５】図１の油圧駆動装置に使用される圧力補償弁の断面構造図（概念図）を示す。
【図６】図２（ａ）とはさらに異なる実施形態である方向制御弁の下流側に圧力補償弁を配置した油圧回路図を示す。
【図７】（ａ）は図２（ａ）とは別の実施形態である２個の走行モータを同期して駆動する油圧駆動装置の部分油圧回路図、（ｂ）は図２（ａ）とはさらに別の実施形態である大きく異なる負荷を有する２個のアクチュエータを駆動する油圧駆動装置の部分油圧回路図、をそれぞれ示す。
【符号の説明】
１原動機２可変ポンプ
４、４’、１４、２８、２９、３０、３６、４４、４８圧力補償弁
８、１８方向制御弁１０、１１、１５、１５’、２０、２５、５０、５１、油圧アクチュエータ
１３シャトル弁１６最高負荷圧力ライン
１７ポンプ容積変更手段
１９、４６スプリング３１差圧制御弁
３３、３５パイロット油路３８、４５ポンプ流量制御弁[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a plurality of directional control valves having a flow rate adjusting function capable of controlling pressure oil flowing into each of a plurality of actuators from discharge oil of one hydraulic pump used in construction machines and the like, The present invention relates to a hydraulic drive device having a plurality of pressure compensating valves for performing respective pressure compensations.
[0002]
[Prior art]
This type of hydraulic drive device is mainly used for construction machines and agricultural machines, and has a load sensing function of controlling a variable pump discharge amount according to a load pressure. Also, when driving a plurality of actuators at the same time, a pump compensating valve is provided in a circuit to each actuator so that a speed change of the actuators does not occur due to a difference between load pressures of the actuators. Are diverted. Further, when the pump discharge amount falls below a predetermined required flow rate of a plurality of drive actuators, a device having a function of distributing the pump discharge amount to each actuator at an appropriate ratio, a so-called anti-saturation function, is also used. For example, in Japanese Patent Publication No. 4-48967 having such an anti-saturation function, the load sensing function applies a spring force and the maximum load pressure of each actuator in a direction to increase the variable pump discharge amount. A pump flow control unit is provided which applies a discharge pressure in a direction to decrease the variable pump discharge amount in opposition to the acting force, and controls the pump discharge amount according to the load pressure. On the other hand, as described in, for example, Japanese Patent Application Laid-Open No. 19409/1992, a pressure compensating valve is disposed downstream of a directional control valve having a flow rate adjusting function, and the pressure compensating valve is opened in each control oil chamber of the pressure compensating valve. An apparatus having an anti-saturation function is disclosed in which the pressure (Pd ') on the upstream side of the pressure compensating valve is applied, and the maximum load pressure (Pm) of the actuator is applied in the direction in which the pressure compensating valve is throttled. .
In Japanese Utility Model Laid-Open No. 4-119604, the pressure (Pz) on the downstream side of the pressure compensating valve is applied to each control oil chamber of the pressure compensating valve in the direction in which the pressure compensating valve is closed so that the variable pump discharge pressure (Pd) is reduced. A differential pressure control valve for generating a secondary pressure (Pc = Pd-Pm) corresponding to the differential pressure from the maximum load pressure (Pm) of the actuator is provided, and the differential pressure control valve outputs the pressure in a direction to open the pressure compensating valve. There is also described a configuration in which a secondary pressure (Pc) and an actuator load pressure (PL) that is a downstream pressure of the direction control valve are applied. Further, FIG. 6 of JP-A-4-54303 discloses a differential pressure detector for detecting the variable pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator, and a control signal receiving the output of the differential pressure detector. And an apparatus that secures an anti-saturation function by a secondary pressure (Pc) output by a proportional control valve that is operated by a control signal of the control apparatus.
[0003]
The flow characteristics of the flow compensating region of such a conventional pressure compensating valve having an anti-saturation function have the following characteristics: an actuator load pressure (PL) and an upstream pressure (Pz) of the directional control valve due to the balance between the open side and the closed side. , Ie, the relationship between the differential pressure before and after the directional control valve (ΔP: hereinafter referred to as the directional control valve differential pressure)
ΔP = Pz−PL = Pd−Pm = Pc
And the directional control valve differential pressure is set to a pressure proportional to the differential pressure between the variable pump discharge pressure and the maximum load pressure of the actuator, that is, the secondary pressure.
However, it has been known that in a hydraulic drive device that performs such pressure compensation, when a load having a large inertia is moved, the system is not stable and the operation is accompanied by hunting. For example, when moving a hydraulic excavator of a construction machine, when moving a large load cylinder such as a swing motor, a traveling motor, or a boom cylinder with a large load, hunting is remarkably exhibited, and operability is impaired. There was a problem. That is, consider a case in which the lever of the direction control valve is moved stepwise by a fixed amount to move an actuator having a large inertia, that is, a turning motor or the like. First, the throttle of the directional control valve is opened and oil flows in, but since the inertia of the actuator is large, the actuator does not move suddenly, so that the flow rate does not flow and the load pressure rises instantaneously. When the load pressure rises, the load pressure acts on the pressure compensating valve and acts to open the pressure compensating valve greatly. Therefore, a large flow rate is supplied, and the actuator is rapidly accelerated. However, once the actuator starts moving, the supplied flow rate is finite, so that the speed increases but the acceleration gradually decreases. Therefore, the load pressure that has rapidly risen gradually decreases as the acceleration decreases, so that the opening of the pressure compensating valve gradually decreases, and the supplied flow rate decreases. On the other hand, when the actuator loses its acceleration and reaches a constant speed, the speed is accelerated by a large initial acceleration, so that the speed is considerably higher than the target speed, and the acceleration of the load pressure at that time is stopped. After that, since the pressure compensating valve has become considerably small, the opening degree of the pressure compensating valve is now smaller, and therefore the directional control valve differential pressure is also small. This will cause the flow rate to decrease and the actuator to begin to decelerate, but due to the large inertia, the actuator will attempt to maintain its speed, causing the load pressure to decrease more rapidly. Therefore, the opening of the pressure compensating valve becomes smaller and the speed of the actuator decreases more.However, when the speed decreases to some extent, the load pressure gradually recovers, and therefore, the opening of the pressure compensating valve gradually increases. In the meantime, the deceleration of the actuator stops and reaches a constant speed.However, since the speed suddenly decreases at the beginning of the deceleration, the speed is considerably lower than the target speed, and at the same time, the load pressure at that point stops. Since the pressure has been restored to a large value, the opening of the pressure compensating valve has also increased, and the differential pressure of the directional control valve has also increased. become.
[0004]
Once the actuator begins to accelerate, it will be in the initial state described above. Thus, the phenomenon of rapid acceleration and rapid deceleration does not readily converge, and hunting continues. Actually, the response delay on the pump device side is further added, so that a more complicated aspect is exhibited. In such a circuit that performs pressure compensation, there is a problem that when a load having a large inertia is moved, the system is not stable and hunting is likely to occur. However, in the aforementioned Japanese Patent Publication No. 4-48967 and Japanese Utility Model Laid-Open No. 4-119604, the pressures acting on the control chambers of the pressure compensating valve in the opening and closing directions are made equal, and the anti-saturation function is also reduced. However, there is no disclosure or suggestion about prevention of hunting. Japanese Patent Application Laid-Open No. 4-54303 discloses that when a plurality of actuators are simultaneously operated, hunting and instability of the pressure compensating valve caused by a decrease in the flow rate due to the flow force generated by the throttle portion of the pressure compensating valve can be reduced in a very small pressure receiving area. The discharge pressure of the pump acts in the direction in which the pressure compensating valve opens, and when the differential pressure between the discharge pressure of the pump and the load pressure of the actuator increases, the outlet flow rate of the pressure compensating valve increases to cancel the flow force and flow. Although there is disclosed an outlet flow which is not affected by force, there is no disclosure or suggestion about prevention of hunting for a load having a large inertia.
Further, in each of the prior arts described above, the maximum load pressure (Pm) is caused to act in a direction to close the pump flow control valve for changing the displacement of the pump from the valve device via a long and narrow pilot oil passage. Therefore, at low temperatures, the viscosity of the hydraulic oil increases, and if the pressure loss in the pipeline from the pump to the valve device becomes excessive, the pressure on the upstream side of the pressure compensating valve in the valve device decreases by the pressure loss. As a result, the directional control valve differential pressure was reduced, and the flow rate supplied to the actuator was significantly reduced.
Also, at least two actuators of a plurality of actuators need to be driven synchronously regardless of the load pressure of the actuators, such as two traveling motor drives that rotationally drive a pair of crawlers of a hydraulic traveling vehicle. At the same time, if each directional control valve is switched by the same stroke, the directional control valve differential pressure before and after each directional control valve is controlled by the pressure compensating valve so that the same flow rate is supplied to each traveling motor. In other words, the hydraulic traveling vehicle should be able to travel straight ahead, but if there is a processing error in the spool of each directional control valve, even if the directional control valve differential pressure is equal, the throttle opening of each directional control valve will differ. Therefore, the flow rate supplied to each traveling motor is not equal, and if there is an error in the pressure receiving area due to the processing error of the pressure compensating valve, the throttle opening of the directional control valve is equal. Also, since not equal the direction control valve differential pressure, there is a problem that the straight running of the hydraulic traveling vehicle can not.
[0005]
Furthermore, when simultaneously operating at least two hydraulic actuators having greatly different loads, such as a hydraulic motor for turning a hydraulic excavator and a hydraulic boom cylinder, the inertia load of the high-load-side actuator is excessive in the initial stage of the simultaneous operation. As a result, excessive pressure is generated in the inflow-side actuator port, and most of the pressure oil flows out of the overload relief valve installed in the inflow-side actuator port into the tank. There is a problem that the driving speed of the boom cylinder, which is a load-side hydraulic actuator, becomes extremely slow, and at the same time, the energy loss of the prime mover due to the relief of the pressure oil flowing into the swing motor increases. Thereafter, when the rotation of the swing motor is completed and the vehicle turns at a steady speed, the load pressure of the swing motor rapidly decreases. The pressure compensating valve for the swing motor was almost fully opened in the initial stage due to excessive swing load pressure, but its opening rapidly decreases as the load pressure suddenly decreases. Therefore, the swing motor decelerates with a shock. There is also a problem that the boom accelerates in the opposite direction due to a margin in the pump (effective) discharge amount due to the deceleration, resulting in a jerky movement.
[0006]
[Problems to be solved by the invention]
Of the present inventionClaims 1 to 14The object of the invention described in (1) above is in view of such a problem in the related art, and the object thereof is that even if it is a single operation or a combined operation, hunting with good operability on both the low load side and the high load side does not occur. It is to provide a hydraulic drive having a stable pressure compensating valve. Another object of the present invention is to provide a hydraulic drive device having a pressure compensating valve which has a simple structure, is low in cost, has high reliability, and can flexibly respond to load conditions. It is an object of the present invention as set forth in claim 1 that the viscosity of hydraulic oil becomes high at low temperatures, the pressure loss in the pipeline from the pump to the valve device becomes excessive, and the supply flow rate to the actuator becomes large. It is an object of the present invention to provide a hydraulic drive device that prevents the reduction.
Of the present inventionClaim 5 and 12 DescriptionSUMMARY OF THE INVENTION An object of the present invention is to provide two traveling motors for rotationally driving a pair of crawlers of a hydraulic traveling vehicle, in which at least two of the plurality of actuators need to be driven synchronously regardless of the load pressure of the actuators. Therefore, even when there is an error in the pressure receiving area due to a processing error of the spool of each direction control valve and a processing error of the pressure compensating valve, a hydraulic control device having a pressure compensating valve capable of obtaining straight traveling of the hydraulic traveling vehicle is provided. To provide.
Of the present inventionClaims 6, 7, 13 as well as 14 Described inAn object of the present invention is to provide a pressure oil that can sufficiently supply pressure oil to a small load side even when actuators having extremely different load magnitudes are simultaneously operated, and a load pressure of an actuator having an extremely large load characteristic. Provided is a hydraulic drive device having a pressure compensating valve that enables smooth operation without shock without a sudden change in the speed of each actuator even when the pressure drops suddenly, and can reduce energy loss and load on the prime mover. Is to do.
[0007]
[Means for Solving the Problems]
For this purpose, according to the first invention of the present invention,Claims 2 and 10A variable pump, a plurality of hydraulic actuators driven by the discharge oil of the variable pump, and a plurality of flow control functions capable of controlling the pressure oil flowing into each of the plurality of actuators. A directional control valve and a plurality of pressure compensating valves for respectively compensating the pressure of each directional control valve, wherein each of the pressure compensating valves is a pressure (Pz) downstream of the pressure compensating valve in a closing direction and the plurality of actuators. Of the pump discharge pressure (Pd), which is the pressure on the upstream side of the pressure compensating valve in the opening direction of the pressure compensating valve, and the actuator load pressure, which is the downstream pressure of the directional control valve. (PL) to act on each other to perform the pressure compensation, and to provide a pump flow regulating valve for communicating the discharge oil of the variable pump with the displacement changing means of the variable pump. A high load pressure (Pm) is applied through the oil passage in a direction to close the pump flow control valve and increase the displacement of the variable pump together with the acting force of the spring of the pump flow control valve, and the pump discharge pressure (Pd) Through a separate oil passage to open the pump flow control valve to act in a direction to reduce the displacement of the variable pump,
At least one of the pressure compensating valves reduces a flow rate of at least one of the pressure compensating valves to the actuator in accordance with an increase in a load pressure of the actuator itself. By doing so, the above problem was solved.
[0008]
【The invention's effect】
TakeClaims 2 and 10According to the configuration, at least one of the pressure compensating valves is configured to reduce the flow rate to the actuator in accordance with the increase of its own load pressure, that is, to reduce the directional control valve differential pressure. Even if there is a sudden change in the load pressure, the actuator load pressure is attenuated and the hydraulic control system is stabilized, and the pressure compensation characteristics of the pressure compensation valve which is not affected by the maximum load pressure of the actuator or the discharge pressure of the variable pump are obtained. Regardless of the single operation or the composite operation, both the low-load side and the high-load side have an unprecedented excellent effect of obtaining stable operability without hunting. Furthermore, the pressure compensation characteristic of the pressure compensation valve is a so-called right-downward characteristic that indicates the degree to which the flow rate of the pressure compensation valve is reduced in accordance with the increase of its own load pressure only by changing the components inside the pressure compensation valve. Can be easily set, and a right-down characteristic that does not cause hunting according to the load characteristic of the actuator itself can be obtained. Further, since the structure of the pressure compensating valve itself is simple, strict accuracy is not required, and the cost can be reduced and high reliability can be obtained.
[0009]
[Means for Solving the Problems]
According to a second aspect of the present invention,Claim 1As in the invention, a variable pump, a plurality of hydraulic actuators driven by discharge oil of the variable pump, and a plurality of directions having a flow rate adjusting function capable of controlling pressure oil flowing into each of the plurality of actuators A plurality of pressure compensating valves for compensating respective pressures of the control valve and each directional control valve; and a secondary pressure (Pc = Pc = equal to a differential pressure between the variable pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator). Pd-Pm), and a pump flow regulating valve for communicating the discharge oil of the variable pump with the displacement changing means of the variable pump. Each of the pressure compensating valves is a pressure compensating valve. The pressure (Pz) on the downstream side of the pressure compensating valve acts in the direction of closing the valve, and the secondary pressure (Pc) output from the differential pressure control valve and the downstream pressure of the directional control valve in the direction of opening the pressure compensating valve. In the hydraulic drive apparatus so as to the pressure compensation that the actuator load pressure (PL) by exerting respectively,
The acting force of the spring of the pump flow control valve acts in a direction to close the pump flow control valve and increase the displacement of the variable pump, and the secondary pressure (Pc) is applied to the pump of the variable pump via an oil passage. The above object has been attained by providing a hydraulic drive device wherein the flow control valve is opened to act in a direction to reduce the displacement of the variable pump.
[0010]
【The invention's effect】
TakeClaim 1According to the configuration of the present invention, the secondary pressure (Pc) is caused to act in the direction of opening the pump flow control valve of the variable pump through the oil passage to reduce the displacement of the variable pump. A load pressure (Pm) is acted on from a valve device through a long and narrow pilot oil passage in a direction to close a pump flow regulating valve for changing the displacement of the pump, and a pump discharge pressure (Pd) is applied through another oil passage to the pump. Compared with the valve in which the flow control valve is operated in the opening direction, the valve according to the second embodiment of the present invention increases the viscosity of the hydraulic oil at a low temperature and causes an excessive pressure loss in a pipe from the pump to the valve device. In this case, the secondary pressure (Pc) in the oil passage generates a secondary pressure (Pc) equal to the differential pressure between the pump discharge pressure (Pd) and the maximum load pressure (Pm) in the valve device. And Therefore, regardless of the magnitude of the pressure loss in the pump discharge pipe, the pump discharge pressure (Pd) of the pump discharge oil passage in the valve device is affected by the acting force of the spring of the pump flow regulating valve with respect to the maximum load pressure (Pm). Therefore, unlike the conventional apparatus, the flow rate does not greatly decrease and the speed of the actuator does not decrease even at a low temperature.
[0011]
More preferably, claims 3, 4, 8 to 9As in the invention described in claim 2, in the invention according to claim 2, at least one of the pressure compensating valves is configured to reduce a flow rate to the actuator in accordance with an increase in a load pressure of the actuator itself, The pressure compensating valve is provided upstream of each of the corresponding directional control valves, and the pressure compensating valve acts on the first pressure receiving area of the first oil chamber in the direction in which the downstream outlet pressure of the first oil chamber closes the valve. Causing the secondary pressure to act on the second pressure receiving area of the second oil chamber in a direction to open the valve, and opening the valve by applying the load pressure of each actuator to the third pressure receiving area of the third oil chamber. And the second and third pressure receiving areas are made substantially the same, and the first pressure receiving area is made larger than the third pressure receiving area.
[0012]
Claim 5 and 12 DescriptionIn the two traveling motors for rotationally driving the pair of crawlers of the hydraulic traveling vehicle, at least two of the plurality of actuators need to be driven synchronously regardless of the load pressure of the actuators. Therefore, it is preferable that values obtained by dividing the third pressure receiving area of the two pressure compensating valves communicating with the two actuators by the first pressure receiving area are equal to each other. As a result, when the supply flow rate to the left and right traveling motors changes and the rotation speed becomes different, the load pressure of the larger flow rate increases, but the flow rate of the pressure compensating valve decreases as the load pressure increases. And the flow characteristics of the pressure compensating valves of the pair of left and right traveling motors are made equal to each other, so that even if there is an error in the pressure receiving area due to the processing error of the spool of the directional control valve and the processing error of the pressure compensating valve. Such errors increase the load pressure of one of the larger flow rates. Since the pressure difference between the discharge pressure and the maximum load pressure is constant, the pressure compensating valve with the larger flow rate works in the direction to reduce the flow rate by decreasing the directional control valve differential pressure due to the increase in the load pressure. The amount decreases, and the rotation of the traveling motor with the higher flow rate decreases. In the other traveling motor, the load pressure and the differential pressure between the discharge pressure and the maximum load pressure do not change, so that the flow rate does not change and the rotation speed does not change, so that good running straightness can be ensured. . On the other hand, when changing direction, the load pressure of the traveling motor with the larger flow rate also rises and works so as to maintain straightness.However, when changing direction, the opening of the direction control valve differs greatly between the left and right, so correction is made. The flow rate is supplied to each traveling motor in accordance with the operation stroke of each directional control valve, and the direction can be changed without maintaining the linearity. Since the present invention does not require a special attached valve in addition to the improvement of the original pressure compensating valve, the present invention has an excellent effect that the overall size of the valve is not increased, the cost is low, and the usability is good. Preferably, if the decrease rate of the flow rate is increased, it is likely to be overcorrected when straight ahead and meander easily, and it is difficult to operate while trying to maintain straightness when turning, and conversely if the decrease rate of the flow rate is small. Since the straightness cannot be ensured because the correction cannot be performed, the value obtained by dividing the third pressure receiving area of the pressure compensating valve by the first pressure receiving area is 0.99 to 0.95 (99 to 95% ) Is desirable.
[0013]
More preferably,Claim 3As in the invention of claim 2, in the hydraulic drive device according to claim 2, at least one of the pressure compensating valves is provided upstream of each corresponding directional control valve, and the pressure compensating valve is provided in the first oil chamber of the first oil chamber. A second pressure receiving area of a second oil chamber, the secondary pressure acting on the second pressure receiving area in a direction in which the valve is opened, and a third pressure being applied to the second pressure receiving area of the second oil chamber. The load pressure of each of the actuators is caused to act on the third pressure receiving area of the oil chamber in a direction in which the valve is opened, and the second and third pressure receiving areas are made substantially the same, and the first pressure receiving area is set to the first pressure receiving area. 3. The hydraulic drive device of claim 3, wherein a load pressure of a first actuator of at least two of the plurality of hydraulic actuators is significantly greater than a load pressure of the other second actuator. In the case of a high load characteristic, the value obtained by dividing the third pressure receiving area of the high load side pressure compensating valve communicating with the high load side actuator by the first pressure receiving area is equal to the low load side pressure compensating valve communicating with the low load side actuator. The value was smaller than a value obtained by dividing the third pressure receiving area of the valve by the first pressure receiving area. With this configuration, when the load pressure of the high-load-side actuator suddenly increases, the flow rate of the high-load-characteristic-side actuator decreases, and the reduced flow rate is supplied to the low-load-side actuator. A decrease in the speed of the actuator can be prevented. Preferably, a value obtained by dividing a third pressure receiving area of the pressure compensating valve of the low load side actuator by a first pressure receiving area is 1 to 0.98, and a third pressure receiving pressure of the pressure compensating valve of the high load characteristic side actuator. It is desirable that the value obtained by dividing the area by the first pressure receiving area is 0.97 to 0.94. There is no change in the actual content of the description.
[0016]
BEST MODE FOR CARRYING OUT THE INVENTION
A hydraulic circuit diagram of a hydraulic drive device according to a first embodiment of the present invention will be described with reference to FIG. A plurality of pressure compensating valves 41 and 42 (only two of which are shown) are connected in parallel to discharge oil passages 23 and 3 of a variable displacement hydraulic pump 2 (hereinafter referred to as a pump) driven by a prime mover 1 such as an engine. Pressure oil flowing into each of the plurality of actuators 10 and 20 (only two of which are shown) can be controlled via the check valve 40 to the output oil passage 6 of each pressure compensating valve for compensating the pressure of each directional control valve. Directional control valves 8 and 18 (only two of them are shown) having flow control functions, and the output sides of these directional control valves are connected to actuators 10 and 20, respectively. The return oil is returned to the tank 12 via the respective directional control valves 8 and 18 again. The load pressure taken out from the actuator load pressure take-out port 7 of the directional control valves 8 and 18 via the load pressure take-out line 9 is the highest load pressure of the actuators 10 and 20 by the shuttle valve 13 (hereinafter referred to as the highest load pressure) ( Pm) is selected. The pressure compensating valves 41 and 42 apply the downstream pressure (Pz) and the maximum load pressure (Pm) of the pressure compensating valves to the control oil chambers of the respective pressure compensating valves in the closing direction, respectively. , 42, the pump discharge pressure (Pd), which is the pressure upstream of the pressure compensating valve, and the actuator load pressure (PL), which is the downstream pressure of the directional control valve, act on the control oil chamber of each pressure compensating valve. When the discharge amount of the pump 2 falls below a predetermined required amount of the actuators 10 and 20, the pressure compensating valves 41 and 42 provide an anti-surgical device that distributes the discharge amount of the pump 2 to the actuator at an appropriate ratio. Has a curation function.
[0017]
Further, a pump flow regulating valve 45 for communicating the discharge oil of the variable pump 2 with the displacement changing means 17 of the variable pump is provided, and the maximum load pressure (Pm) can be varied by closing the pump flow regulating valve 45 via the oil passage 35. Acts in the direction of increasing the displacement of the pump 2 together with the acting force of the spring 46 of the pump flow control valve, and the pump discharge pressure (Pd) is opened via another oil passage 23 ′ to open the pump flow control valve 45 and the variable pump. By acting in the direction of decreasing the displacement of 2 and balancing the pump discharge pressure Pd with the maximum load pressure Pm and the acting force set in advance by the spring 46, the acting force of the pump discharge pressure Pd becomes If it is larger than the maximum load pressure Pm and the acting force of the spring 46, the pump discharge is performed so as to reduce the displacement of the variable pump 2. When the acting force of the force Pd is smaller than the acting force of the maximum load pressure Pm and the spring 46, the displacement of the pump 2 is controlled to be large, whereby the pump 2 is controlled in accordance with the (maximum) load pressure. It has a load sensing function for controlling the discharge amount of the variable pump 2. According to the first invention of the present invention, in the hydraulic drive device shown in FIG. 1, the pressure compensating valves 41 and 42 communicating with the respective actuators in accordance with the increase of the load pressure of the respective actuators.At least oneThe flow rate is reduced.
[0018]
With such a configuration, in the first invention of the present invention, the pressure compensating valves 41 and 42At least oneThe pressure receiving area of the closing direction oil chamber is made larger than that of the opening direction oil chamber, and the flow rate of the pressure compensating valve connected to the actuator is reduced in accordance with the increase of the actuator load pressure, that is, the directional control valve differential pressure is reduced. Therefore, even if there is a sudden change in the load pressure of the actuator, the load pressure of the actuator is attenuated, the hydraulic control system is stabilized, and the pressure compensating valve is not affected by the maximum load pressure of the actuator or the discharge pressure of the variable pump. Pressure compensation characteristics can be obtained, and both single operation and combined operation can provide stable operability without hunting on both the low load side and the high load side. became.
[0019]
A hydraulic circuit diagram of a hydraulic drive device according to a second embodiment of the present invention will be described with reference to FIG. The same parts as those in the embodiment of FIG. 1 are denoted by the same reference numerals, and a part of the description is omitted. In the hydraulic circuit diagram of FIG. 2A, the shuttle valve 13 selects the maximum load pressure (Pm) of the actuators 10 and 20, and the differential pressure between the variable pump discharge pressure (Pd) and the maximum load pressure (Pm). ToequalA differential pressure control valve 31 for generating a secondary pressure (Pc) is provided in the valve device 22, and each of the pressure compensating valves 4, 14 is connected to the outlet pressure (6) of the downstream side 6 of the pressure compensating valve in a direction to close the pressure compensating valve. Pz) is actuated to open the pressure compensating valves 4 and 14 from the differential pressure control valve 31 in the direction of opening the secondary pressure (Pc) of the secondary pressure line 32 and the actuators 10 and 20 which are downstream pressures of the directional control valves. The load pressure (PL) of the load pressure line 34 taken out of the above is applied. Further, a pump flow regulating valve 38 is provided for communicating the discharge oil of the variable pump 2 with the displacement changing means 17 of the variable pump, and the action force of the spring 19 of the pump flow regulating valve is closed by closing the pump flow regulating valve and displacing the pump 2. The secondary pressure Pc is caused to increase by opening the pump flow control valve 38 via the oil passage 33 so as to reduce the displacement of the pump 2. The secondary pressure Pc is preset by the secondary pressure Pc and the spring 19. When the force exerted by the secondary pressure Pc is greater than the force exerted by the spring 19 by balancing the applied force, the secondary pressure Pc is adjusted by the spring to reduce the displacement of the variable pump 2. If the acting force is smaller than the acting force of the pump 19, the load sensing device that operates the volume changing means 17 so as to increase the displacement of the pump 2 Having.
[0020]
Here, the operation of the hydraulic drive device of FIG. 2A will be described. The pressure on the upstream side 6 of the directional control valves 8 and 18 in the respective pressure compensating valves 4 and 14 increases the load pressure on the respective actuators on the downstream side. Since the respective pressure compensating valves 4 and 14 act so as to balance with the sum of (PL) and the secondary pressure (Pc), if the respective directional control valve differential pressures have the same pressure receiving area, It is equal to the above-mentioned secondary pressure (Pc) regardless of the load pressure of the actuator. That is, it is equal to the differential pressure between the pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator. Further, the secondary pressure (Pc) is guided to the pump flow control valve 38 through the oil passage 33 and is balanced with the acting force of the spring 19 of the pump flow control valve 38, so that the discharge pressure (Pd) of the pump 2 Is controlled so that the secondary pressure (Pc) becomes equal to the pressure corresponding to the acting force of the spring 19. This means that the variable pump discharge pressure (Pd) is controlled so that the secondary pressure (Pc) and the pressure corresponding to the acting force of the spring 19 become equal. Therefore, the directional control valve differential pressure of each directional control valve 8, 18 is also controlled to a pressure corresponding to the acting force of the spring 19. With such a configuration, if the pump discharge amount becomes insufficient, the differential pressure between the pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator is obtained.be equivalent toSince the secondary pressure (Pc) cannot secure the differential pressure set by the spring 19, the respective directional control valve differential pressures are also smaller than the set values. Are equal to each other, the flow to the respective actuators 10 and 20 is diverted to a flow equal to the ratio of the throttle opening of the directional control valves 8 and 18, and thus has an anti-saturation function.
[0021]
With such a configuration, in the second aspect of the present invention, the secondary pressure (Pc) is reduced in the direction in which the pump flow control valve 38 of the variable pump 2 is closed via the pilot oil passage 33 and the displacement of the variable pump 2 is reduced. 1, the maximum load pressure Pm is applied from the valve device 43 to the direction in which the pump flow regulating valve 45 is closed via the long and narrow pilot oil passage 35, as shown in FIG. Compared with the case where the discharge pressure Pd acts in the direction in which the pump flow regulating valve 45 is opened, the viscosity of the hydraulic oil at the time of low temperature increases in the second invention of the present invention, and the pipe line 23 from the pump 2 to the valve device 22 increases. Even if the pressure loss in the oil passage 32 becomes excessive, the secondary pressure (Pc) of the oil passage 32 is determined by the difference between the pump discharge passage 3 in the valve device 22 and the maximum load pressure (Pm). Since the pressure corresponding to (Pc) is generated, regardless of the magnitude of the pressure loss in the pump discharge pipe 23 from the pump 2 to the valve device 22, the pump discharge from the pump discharge oil passage 3 in the valve device 22 is performed. Since the pressure (Pd) is controlled to a pressure corresponding to the acting force of the spring 19 with respect to the maximum load pressure of the actuator, even at a low temperature, unlike the conventional device, the flow rate does not decrease significantly, The speed of the actuator does not decrease.
[0022]
The hydraulic circuit of the embodiment shown in FIG. 2A is also disclosed in the first invention of the present invention.ToPressure compensating valve4, 14The pressure receiving area of the closing direction oil chamber can be made larger than that of the opening direction oil chamber, and the flow rate of the pressure compensating valve communicating with the actuator can be reduced in accordance with an increase in the actuator load pressure. As a result, even if there is a sudden change in the self-load pressure, the actuator load pressure is attenuated, the hydraulic control system is stabilized, and the pressure compensation characteristic of the pressure compensation valve is not affected by the maximum load pressure of the actuator or the discharge pressure of the variable pump. Irrespective of the single operation or the combined operation, both the low-load side and the high-load side have an unprecedented excellent effect of obtaining stable operability without hunting.
[0023]
Referring to FIG.Of FIG.A sectional structural view of an embodiment of the pressure compensating valves 4 and 14 used in the hydraulic drive device is shown. Since the cross-sectional structure of the pressure compensating valve 4 is the same as that of the pressure compensating valve 14, the pressure compensating valve 4 is representatively shown here. However, as described below, the pressure compensating valves 4 and 14 may have different pressure receiving areas of the respective pressure chambers. The pressure compensating valve 4 is slidable in the main body 101, the main body hole 128 having two inner diameters of the main body 101, the small-diameter main body hole 111 and the large-diameter main body hole 130, and the small-diameter main body hole 111 (inner diameter d 3). And a spool 112 having first and second large-diameter lands 133 and 134 slidably fitted with the small-diameter portion 132 and the large-diameter main body hole 130 (inner diameter d2). The actuator 101 has a load pressure port 103, a secondary pressure port 104, an outlet port 105, an inlet port 102 communicating with a pump discharge oil passage, and a tank port 106 which are sequentially provided in the actuator 101. A small-diameter portion 132 provided at one end of the spool 112 fitted into the small-diameter main body hole 111 forms a third oil chamber 119 communicating with the load pressure port 103 so as to be able to abut on the main body hole end surface 127 via the spring 118 and the spool. The other end 114 of 112 forms an oil chamber 124 that communicates with the tank port 106.
[0024]
A second oil chamber 113 communicating with the secondary pressure port 104 is formed in a large-diameter main body hole 130 surrounding a joint between the small-diameter portion 132 of the spool 112 and the first large-diameter land 133, and the other end 114 of the spool 112 is formed. A piston 117 is slidably inserted in an oil-tight manner into a bore 116 (inner diameter d1) provided in the tank, and the other end of the piston 117 is made to be able to abut on the other body hole end surface 126 to form a tank. In tank oil chamber 124 that leads to port 106. A first oil chamber 121 communicating with the outlet port 105 via the pilot oil path 123 is formed between the spool 112 and the piston 117 in the axial hole 116, and a first pressure receiving pressure of the first oil chamber 121 is formed. The area A1 is determined by the cross-sectional area of the piston 117, the second pressure receiving area A2 of the second oil chamber 113 is determined by the area obtained by subtracting the cross-sectional area of the small-diameter hole 111 from the cross-sectional area of the large-diameter main body hole 130, and the third oil chamber 119. The third pressure receiving area A3 is formed by the small-diameter portion 132 in cross-sectional area, and the spool 112 is provided with an outlet port 105 provided on the second large-diameter land 134 facing the first large-diameter land portion 133. And an openable / closable throttle 115 for narrowing the space between the inlet ports 102. In the first oil chamber 121 communicating with the outlet port 105, the outlet pressure Pz acts on the spool 112 in the direction to close the throttle portion 115 to the left as viewed in the figure, and the second pressure receiving area A2 of the second oil chamber 113 , The secondary pressure Pc acts on the spool 112 in a direction to open the throttle portion 115 rightward as viewed in the figure, and the load pressure PL applies the spool 112 to the third pressure receiving area A3 of the third oil chamber 119. It acts in a direction to open the throttle portion 115 rightward as viewed in the drawing.
[0025]
In the embodiment of FIG. 3, the third pressure receiving area A3 and the second pressure receiving area A2 are made substantially equal, and the outer diameter d3 of the spool 112 small diameter portion 132 is slightly smaller than the outer diameter d1 of the piston 117 (d3 <D1) The third pressure receiving area A3 is smaller than the first pressure receiving area A1. In addition, when the spool 112 makes a maximum stroke in the left direction as viewed in FIG. 3, the left end surface of the spool comes into contact with the left end surface 127 of the main body of the third oil chamber 119, and the throttle unit 115 is closed. I have. Conversely, when the maximum stroke is made in the right direction, the right end face 114 of the spool and the right end face of the piston 117 abut on the right end face 126 of the main body hole, and the throttle portion 115 is fully opened. In the middle stroke of the spool 112, the opening degree is increased proportionally to the stroke amount of the spool to the right by the throttle portion 115 of the spool. The spring 118 is for moving the spool 112 to the right when the directional control valves 8 and 18 are not operated to keep the throttle portion 115 open, and its acting force is extremely weak. . FIG. 3 is for conceptually showing the principle of operation. Although both ends of the main body hole 128 are not open, the main body hole is actually a stepped through hole (not shown) or a machined hole from the right side. It can be structured to be closed by a method such as a screw plug (not shown).
[0026]
Next, the operation of the embodiment of FIG. 3 will be described. Consider the balance of the forces acting on the spool 112 of the pressure compensating valve. First, the force acting on the spool 112 in the right direction in the figure, that is, the direction in which the throttle unit 115 is opened, is as follows: load pressure PL, pump discharge pressure Pd, maximum load pressure Pm, and secondary pressure Pc (Pc = Pd− Pm)
(A3 · PL) + (A2 · Pc) ... (1)
Conversely, the force acting in the left direction in the figure, that is, the direction in which the restricting portion 115 is closed, is Pz, where the outlet pressure of the directional control valve upstream 6, that is, the outlet port 105 is Pz.
(A1 · Pz) ... (2)
It becomes. Here, when controlling the pressure compensating valve, since the forces in both directions are balanced, the expressions (1) and (2) are equal.
(A3 · PL) + (A2 · Pc) = (A1 · Pz) (3)
The following relationship holds. However, the acting force of the spring 118 is so weak that it is ignored.
If it is assumed here that the outer diameter d3 of the spool small diameter portion is equal to the outer diameter d1 of the piston 117, then A3 = A1. From the equation (3), the directional control valve differential pressure ΔP = (Pz−PL) becomes
ΔP = (Pz−PL) = (A2 / A3) · Pc (4)
It becomes. Therefore, the direction control valve differential pressure ΔP is determined to be a constant value by the secondary pressure Pc, the outer diameter d2 of the spool 112, the outer diameter d3, and the outer diameter d1 of the piston 117. It always has a constant value. In the saturation state, the secondary pressure Pc decreases according to the situation, and the directional control valve differential pressure also decreases. However, as described above, since the differential pressures are equal, the flow rate to the respective actuators 10 and 20 is reduced. Is diverted into a flow rate equal to the ratio of the throttle opening degrees of the directional control valves 8 and 18, and thus has an anti-saturation function. Not affected. Note that the third pressure receiving area A3 and the second pressure receiving area A2 may or may not be equal. If A2 = A3, ΔP = Pc, and if A2 ≠ A3, the absolute value of ΔP can be variously changed according to the ratio of A2 to A3 as shown in equation (4). Note that the first pressure receiving area is determined in relation to the third pressure receiving area.
[0027]
In the first aspect of the present invention, the outer diameter d1 of the piston 117 is slightly larger than the outer diameter d3 of the small diameter portion of the spool 112 (d1> d3). Therefore, when A3 = k · A1 (where k <1) is substituted into equation (3),
k · A1 · PL + A2 · Pc = A1 · Pz (5)
In Equation (5), for convenience, k = {1- (1-k)},
{1- (1-k)} · A1 · PL + A2 · Pc = A1 · Pz
Become, deform,
A1 · PL-A1 · (1-k) · PL + A2 · Pc = A1 · Pz
PL− (1-k) · PL + (A2 / A1) · Pc = Pz
− (1-k) · PL + (A2 / A1) · Pc = Pz−PL
Therefore, the direction control valve differential pressure ΔP is
ΔP = (Pz−PL) = (A2 / A1) · Pc− (1−k) · PL (6)
Or, if A1 = A3 / k is substituted.
ΔP = [(k · A2) / A3] · Pc− (1−k) · PL (7)
Get. Here, since the constant k is a value smaller than 1, the second term on the right side of the equations (6) and (7) is a negative value. According to the equations (6) and (7), the directional control valve differential pressure ΔP is a linear expression of the secondary pressure Pc and the actuator load pressure PL, and the respective directional control valve differentials increase in accordance with the increase in the actuator load pressure PL. The pressure ΔP decreases and the flow rate decreases. That is, a downward pressure compensation characteristic in which the flow rate decreases in accordance with an increase in the actuator load pressure PL is obtained.
This is the case where only one directional control valve is operated, or even when two or more directional control valves are simultaneously operated, the condition that the maximum discharge amount of the pump is equal to or more than the required flow rate of all actuators, that is, This is always true if the saturation state has not been reached. That is, in this state, the secondary pressure Pc is maintained at a constant pressure set by the acting force of the spring 19 as described above. On the other hand, since the load pressure PL is the load pressure of each actuator, the rightward downward pressure that depends only on the load pressure of each actuator consistently regardless of the load pressure of other actuators, the maximum load pressure of the actuator, or the pump discharge pressure. A compensation characteristic is obtained. On the other hand, in the saturation state where the pump discharge amount is insufficient, the secondary pressure Pc becomes a pressure Pc ′ smaller than the acting force set by the spring 19, and the magnitude of the Pc ′ depends on the degree of insufficient flow rate and becomes a constant value. No. However, the same secondary pressure Pc 'acts on all the pressure compensating valves, and all of the discharge amount of the pump is distributed to each actuator at an appropriate ratio.
Further, if only one directional control valve is operated to be in a saturation state, all the discharge amount flows to one actuator regardless of the actuator load pressure PL.
[0028]
Next, a case where two directional control valves are operated to reach a saturation state will be described. For the sake of simplicity, it is assumed that there is no change in the opening of both directional control valves.In this state, only the load pressure of the actuator on one side increases and the load pressure on the other side does not change. I do. The directional control valve differential pressure ΔP on the side where the load pressure has risen decreases as the actuator load pressure PL rises according to the equations (6) and (7). However, since the differential pressure of the first term is a small pressure Pc ', the flow rate itself is small. The directional control valve differential pressure on the other side where the load pressure does not change does not change in the second term of the equations (6) and (7) because the actuator load pressure PL does not change, but the flow rate decrease on the side where the load pressure increases As a result, there is a margin in the flow rate as a whole, and the secondary pressure Pc 'increases, so that the first term increases, the directional control valve differential pressure increases, and the flow rate increases. In other words, the entire pump discharge amount is distributed to the actuator as a whole, and the flow rate on the side where the load pressure does not change increases by the decrease in the flow rate on the side where the load pressure increases. Therefore, in the saturation state, the flow rate on the side where the load pressure does not change will increase due to the load pressure on the side where the load pressure increases, even though the load pressure of the own apparatus is constant, but originally required in the saturation state. This characteristic cannot be a cause of hunting because an overshoot relative to the target speed does not occur because the actual flow rate is insufficient with respect to the flow rate. Rather, the increase in the flow rate on the side where the load pressure does not increase has the advantage that the speed that is closer to the target can be made higher than the insufficient flow rate.
[0029]
On the other hand, when the other load pressure decreases, the opposite phenomenon occurs. That is, the flow rate increases on the side where the load pressure decreases, and the flow rate decreases on the side where the load pressure is constant. Further, when the fluctuation of the load pressure rises or falls at the same rate, the shunt ratio of each flow rate changes without change. This is also true when three or more directional control valves are operated simultaneously. Thus, according to the present invention, stable controllability without hunting can be obtained even when the pump discharge amount is sufficient or in the saturation state.
Further, it goes without saying that the pressure compensation characteristic can be set to an arbitrary value by changing the value of the constant k. That is, the smaller the value of k, the greater the degree of downward pressure to the right of the pressure compensating valve. This means that the right-downward degree can be set according to the load characteristics of each actuator. In addition, the setting can be simply changed only by changing the outer diameter d1 of the piston 117, without changing the main body 101 itself.
Note that the value of the constant k is determined according to the actual device. However, in an actuator that is likely to cause hunting, the hunting is easily performed when the reduction rate of the compensation flow rate is small, and the original flow rate is fixed when the reduction rate is large. Therefore, a pressure compensation function of about 0.99> k> 0.95 (0.99 to 95%) is appropriate. As described above, since not only the degree of downward inclination but also various values of k can be easily obtained with the same main body, various pressure compensating valves according to load conditions can be easily obtained.
[0030]
FIG. 7 (a)As shown in FIG. 2A, in a hydraulic drive device having a circuit shown in FIG. 2A, a plurality of actuators such as two drive motors 15 and 15 'for driving a pair of crawlers of a hydraulic travel vehicle are driven. When at least two actuators 15, 15 'need to be driven synchronously regardless of the load pressure of the actuators, the two pressure compensating valves 28, 29 leading to said two actuators 15, 15' It is desirable that values obtained by dividing the third pressure receiving area A3 by the first pressure receiving area A1 are equal. With such a configuration, when the supply flow rates of the left and right traveling motors 15 and 15 'change and the rotation speeds are different, the load pressure of the supply flow rate which increases is increased. And the flow characteristics of the pressure compensating valves of the pair of left and right traveling motors are made equal to each other, so that the processing error of the spools of the directional control valves 8 and 18 and the pressure receiving pressure due to the processing error of the pressure compensating valves. Even when there is an error in the area, such an error causes the load pressure PL of one of the larger flow rates to increase. Here, since the differential pressure Pc between the discharge pressure Pd and the maximum load pressure Pm is constant, the pressure compensating valve having the larger flow rate decreases the flow rate by decreasing the directional control valve differential pressure by increasing the load pressure. Therefore, the inflow amount decreases, and the rotation of the traveling motor with the higher flow rate decreases. In the other traveling motor, the load pressure and the differential pressure between the discharge pressure and the maximum load pressure do not change, so that the flow rate does not change and the rotation speed does not change, so that good running straightness can be secured. . On the other hand, when changing direction, the load pressure of the traveling motor with the larger flow rate also rises and works so as to maintain straightness.However, when changing direction, the opening of the direction control valve differs greatly between the left and right, so correction is made. The flow rate is supplied to each traveling motor in accordance with the operation stroke of each directional control valve, and the direction can be changed without maintaining the linearity. The present invention does not require a special attached valve in addition to the improvement of the original pressure compensating valve, so that the overall effect of the present invention is that the overall size of the valve is not increased, the cost is low, and at the same time the usability is good.
[0031]
Preferably, if the decrease rate of the flow rate is increased, it is likely to be overcorrected when straight ahead and meander easily, and it is difficult to operate while trying to maintain straightness when turning, and conversely if the decrease rate of the flow rate is small. Since the straightness cannot be ensured because the correction cannot be performed, the value obtained by dividing the third pressure receiving area of the pressure compensating valve by the first pressure receiving area is 0.99 to 0.95 (99 to 95% ) Is desirable.
[0032]
As shown in FIG. 7 (b), more preferably, in the hydraulic drive device according to the second aspect of the present invention, as in the turning hydraulic motor of at least two actuators 11, 25 of the plurality of hydraulic actuators. The load pressure of the high-load-side actuator 25 is extremely higher than the load pressure of the low-load-side actuator 11 such as the other boom hydraulic cylinder.For load characteristics,Preferably, a value obtained by dividing the third pressure receiving area A3 of the high load side pressure compensating valve 36 communicating with the high load side actuator 25 by the first pressure receiving area A1 is a low load side pressure communicating with the low load side actuator 11. The value was obtained by dividing the third pressure receiving area A3 of the compensating valve 30 by the first pressure receiving area A1. With such a configuration, high loadCharacteristicWhen the load pressure of the side actuator 25 suddenly rises,CharacteristicSince the flow rate of the side actuator is reduced and the reduced flow rate is supplied to the low load side actuator 11, it is possible to prevent the speed of the low load side actuator 11 from decreasing. Preferably, the value obtained by dividing the third pressure receiving area A3 of the pressure compensating valve 30 of the low load side actuator 11 by the first pressure receiving area A1 is 1 to 0.98, and the pressure compensating valve 36 of the high load side actuator 25 is set. It is desirable that the value obtained by dividing the third pressure receiving area A3 by the first pressure receiving area A1 is 0.97 to 0.94.
There is no change in the actual content of the description.
[0033]
Further, when the swing load pressure of the high load side actuator 25 is excessive, the opening degree of the pressure compensating valve 36 is reduced and the flow rate supplied to the swing motor 25 side decreases. At the same time as the useless relief flow rate flowing away from the relief valve is reduced, the increase in the load pressure of the swing motor 25 is also suppressed. Therefore, a reduction in the speed of the boom hydraulic cylinder, which is the low-load-side actuator 11, can be prevented by an amount corresponding to a reduction in the useless relief flow rate. Thereafter, the load pressure decreases as the speed of the swing motor 25 increases and the acceleration decreases, so that the opening of the pressure compensating valve 36 gradually increases, and the flow rate gradually increases with the decrease in the swing load pressure. Therefore, a gentle acceleration of the turning motor 25 can be obtained. Further, when the rotation of the swing motor is completed and the swing is performed at a steady speed, the swing load pressure sharply decreases, and the load pressure of the boom cylinder increases. At this time, the opening of the pressure compensating valve 36 on the side of the swing motor 25 is not fully opened, but is in the process of gradually increasing the opening from the state of the throttle, and is still in the state of being throttled. Therefore, even if the load pressure of the swing motor 25 suddenly decreases, the rate at which the opening degree of the pressure compensating valve 36 on the swing motor 25 side sharply decreases decreases, and the swing motor 25 decelerates with a shock. Nothing. Conversely, the opening degree of the pressure compensating valve 30 on the boom cylinder 11 side is relatively large since the secondary pressure is secured to some extent from the initial stage of turning of the turning motor 25. Even if the rotation of the rotation motor is completed and the vehicle turns at a steady speed, the opening degree does not increase rapidly as in the conventional example, and acceleration with a shock does not occur.
[0034]
Referring to FIG.Of FIG.A description will be given of a pressure compensating valve 4 ′ of an embodiment different from FIG. 3 used for the hydraulic drive device. The same parts as those in the embodiment of FIG. 3 are denoted by the same reference numerals, and a part of the description will be omitted. The pressure compensating valve 4 'in FIG. 4 differs from the embodiment in FIG. 3 in the configuration of the pressure receiving areas A3 and A2 acting in the opening direction of the pressure compensating valve. That is, in FIG. 4, the main body 201 has only a large-diameter hole (inner diameter d2) in the main body hole 228, and a spool 212 having first, second, and third large-diameter lands 209, 210, and 211 is provided in the main body hole 228. It is inserted. Then, instead of the small-diameter hole 111 (inner diameter d3) in FIG. 3, an auxiliary piston 217 having an outer diameter of d3 is slidably inserted in an auxiliary axial direction hole 202 provided at an outer end 214 of the spool 212 in a nested manner. Things. It has a secondary pressure port 204, a load pressure port 203 of the actuator, an outlet port 105, an inlet port 102 communicating with the pump discharge oil passage, and a tank port 106 which are sequentially provided in the main body 201 along the main body hole 228. The outer end of the auxiliary piston 217 is made to be able to contact the end face 227 of the main body hole 228 to form a second oil chamber 213 communicating with the secondary pressure port 204. A third oil chamber 220 is formed between the spool 212 and the auxiliary piston 217 in the auxiliary axial hole 202 through a spring 218 and through the auxiliary pilot oil passage 223 to the load pressure port 203. The first pressure receiving area A1 of the first oil chamber 121 is determined by the sectional area of the piston 117, and the second pressure receiving area A2 of the second oil chamber 213 is determined by the sectional area of the auxiliary piston 217 from the sectional area of the main body hole 228. The third pressure receiving area A3 of the third oil chamber 220 is formed by the cross-sectional area of the auxiliary piston 217, respectively, by the subtracted area.
[0035]
According to this configuration, if the relationship between the diameters d1, d2, and d3 is the same as that of the embodiment of FIG. 3, the secondary pressure Pc, that is, the differential pressure between the pump discharge pressure and the maximum load pressure of the actuator is equal to the pump pressure. Since the load pressure PL is sufficiently large with respect to the secondary pressure Pc because of the acting force of the spring 19 of the flow control valve 38, the auxiliary piston is pressed against the left end face of the main body hole and is similar to the embodiment shown in FIG. Action is obtained.
[0036]
In the embodiment of FIG. 4, when the load pressure PL becomes smaller than the secondary pressure Pc due to a negative load such as a self-propelled load, the auxiliary piston 217 is moved to the left end face of the main body hole. The secondary pressure Pc acts on the pressure receiving area A3 where the load pressure PL acts on the spool 212 so as to bias the spool 212 away from the secondary pressure Pc. In this case, the load pressure is assumed to be the same as Pc and is controlled, and the differential pressure of the directional control valve is slightly increased, so that the flow rate is slightly increased. However, in contrast to the embodiment of FIG. 3 having a stepped hole, the embodiment of FIG. 4 has an advantage that the hole processing does not need to be stepped and the processing is easy, and the secondary pressure Pc is Since it is originally small, there are few practical problems caused by the secondary pressure Pc acting on the outer end face of the auxiliary piston. Further, it is more suitable for a device in which the load pressure itself is always a certain pressure or higher. Become.
[0037]
As described above, the pressure compensating valves 4, 4 'shown in FIGS.And FIG.However, the pressure compensating valves 4 and 4 'used in the hydraulic drive device of the present invention are not shown in FIG.And of FIG.Other than the circuit example, the present invention is applicable. That is, as described above, the load pressure PL and the secondary pressure Pc act in the opening direction, and the upstream pressure Pz (downstream of the pressure compensating valve) Pz of the direction control valve acts in the closing direction to control. What is necessary is just a pressure compensating valve. For example, it can be used for the circuit shown in FIG.
[0038]
Instead of applying the secondary pressure Pc to the pump control circuit shown in FIG. 2A via the pilot oil passage 33, the discharge oil of the variable pump 2 (as shown in FIG. 1) in FIG. And a pump flow regulating valve 45 that communicates with the variable pump displacement volume changing means 17. The maximum load pressure (Pm) is closed via the oil passage 35 and the displacement of the variable pump 2 is increased. The pump discharge pressure (Pd) is actuated in a direction to open the pump flow regulating valve 45 via another oil passage 23 'to reduce the displacement of the variable pump 2, thereby reducing the acting force of the pump discharge pressure Pd. Alternatively, the maximum load pressure Pm may be balanced with the action force set in advance by the spring 46.
[0039]
As shown in FIG. 2 (c), in FIG. 2 (a), the variable pump discharge pressure (Pd) from the pump discharge pressure discharge line 23 'and the maximum load pressure discharge line 16 are used to generate the secondary pressure Pc. The differential pressure from the maximum load pressure (Pm) of the actuator is detected by a differential pressure detector 60, and the output of the differential pressure detector 60 is input to a control device 61 to generate and output a control signal 62, thereby controlling The anti-saturation function may be ensured by the secondary pressure (Pc) output by the electromagnetic proportional valve 63 operated by the signal 62. 64 is a pilot pump.
[0040]
FIG.2A is a hydraulic circuit diagram showing an embodiment of a hydraulic drive device different from that of FIG. 2A, and is similar to Japanese Patent Application Laid-Open No. 4-140909 in that a pressure compensating valve is arranged downstream of a directional control valve. 2A are denoted by the same reference numerals, and a part of the description will be omitted.FIG.As shown in the figure, 50 and 51 are hydraulic actuators which guide the pressure oil in the discharge oil passage 3 through a check valve 40, directional control valves 53 and 54 having a flow control function, and pressure compensating valves 44 and 48, The return oils 50 and 51 are returned from the direction control valves 53 and 54 to the tank T via the tank oil passage 12. The highest load pressure among the load pressures of the actuators is selected by the shuttle valve 13, and both pressure compensating valves 44 and 48 are operated together with the springs 44a and 48a in the closing direction, so that the downstream pressures Pd 'of the direction control valves 53 and 54 are reduced. The pressure compensating valves are operated in the opening directions. The differential pressure across the directional control valves 53 and 54 is configured to match the differential pressure between the variable pump discharge pressure Pd and the maximum load pressure Pm, as in FIG. 2 (a), so that an anti-saturation function is provided. I have. A differential pressure control valve 31 for generating a pressure Pc corresponding to a differential pressure between the pump discharge pressure Pd of the discharge oil passage 3 in the valve device 24 and the maximum load pressure Pm of the outlet oil passage 16 selected by the shuttle valve 13 is shown. 2 (a), the secondary pressure Pc generated by the differential pressure control valve 31 is supplied through the secondary pressure line 32 and the pilot oil passage 33 to the pump flow control valve 38 of the pump device 21. The displacement of the variable pump 2 is opened to actuate the displacement changing means 17 so as to reduce the discharge of the variable pump 2. The spring 19 of the pump flow control valve 38 is closed, the pump flow control valve is closed, and the pump flow control valve is closed. Control is performed to increase the discharge amount.
[0041]
In FIG.Explaining the operation of the hydraulic drive device, the pressure at the discharge port of the variable pump 2 is higher than the pressure of the discharge conduit 3 in the valve device 24 by an amount corresponding to the pressure loss in the discharge conduit 23. The pump discharge pressure Pd of the discharge oil passage 3 is determined only by the maximum load pressure Pm and the acting force of the spring 19, and does not depend on the temperature of the hydraulic oil. That is, the balance formula of the force in the pump flow regulating valve 38 is
Pc = acting force of spring 19
And the secondary pressure Pc is obtained from the differential pressure control valve 31
Since Pc = Pd−Pm,
Pd−Pm = acting force of spring 19
Holds, and the pump discharge pressure becomes
Pd = Pm + acting force of spring 19
It becomes.
Also, from the balance of the forces in the pressure compensating valves 44 and 48,
Pd '= Pm + action force of spring 44a
It becomes.
Therefore, the differential pressure between the directional control valves 53 and 54 = Pd−Pd ′ = the acting force of the spring 19 The acting force of the car spring 44a
The differential pressure before and after the direction control valve is determined only by the acting force of the spring 19 of the pump flow regulating valve 38 and the acting force of the springs 44a, 48a of the pressure compensating valves 44, 48. Since it does not depend on the temperature, it is possible to provide a hydraulic drive device that is not affected by the temperature of the hydraulic oil, as in FIG.
On the other hand, in the case of the saturation state, the relationship of the above expression is not established because the pump discharge amount is insufficient. However, the pressure Pd 'on the downstream side of the direction control valves 53 and 54 is the sum of the maximum load pressure Pm and the acting force of the springs of the pressure compensating valve springs 44a and 48a. The pressure downstream of the control valve will be the same. On the other hand, since the upstream side of all the directional control valves communicates in parallel with the discharge conduit 3, the pressure is the same as Pd. Therefore, the differential pressures before and after all the directional control valves are the same, and the variable pump discharge amount is divided at a rate proportional to the ratio of the opening degrees of the respective directional control valves, as in the case of FIG. It will have a saturation function.
[0042]
Referring to FIG. 5, a pressure compensating valve 41 used in FIG. 1, which is an embodiment of the first invention of the present invention, will be described. The body 301 of the pressure compensating valve 41 is divided into a first body 301a and a second body 301b, which are assembled together by a suitable method such as bolting (not shown), and a small-diameter hole 321 is formed in the first body 301a. And a medium diameter hole 322 following the small diameter hole, a first spool 311 fitted to the small diameter hole 321 and a second spool 312 fitted to the medium diameter hole 322 are arranged, and the second body 301b A large-diameter hole 323 following the large-diameter hole 322 and an auxiliary small-diameter hole 325 having the same diameter as the small-diameter hole 321 following the large-diameter hole are provided, and the third spool 310 fitted into the large-diameter hole 323 and the auxiliary small-diameter hole 325 is large. It has first and second large-diameter lands 313 and 314 fitted in the diameter hole 323 and an auxiliary small-diameter portion 315 fitted in the auxiliary small-diameter hole 325, and has a first spool 311 and a small-diameter hole 321 end face 3. Wherein it is through a spring 350 to press the spools between 0. Further, along the body 301, an auxiliary inlet port 341 communicating with the pump discharge oil passage 3 and communicating with the small diameter hole 321; an actuator load pressure port 342 communicating with the actuator load pressure line 34 and communicating with the medium diameter hole 322; A tank port 343 communicating with a large-diameter hole 323 surrounding a contact portion between the second spool 312 and the third spool 310; an outlet port 344 communicating with a large-diameter hole 323 between the first and second large-diameter lands 313 and 314; An inlet port 345 that communicates with the discharge oil passage 3 and is opened to be throttled by an openable and closable throttle portion 316 provided in the second large-diameter land 314, and communicates with the highest load pressure extraction line 16 of the actuator and is Maximum load pressure passing through the large-diameter hole 323 at the continuous portion between the large-diameter land 314 and the auxiliary small-diameter portion 315 Over door 346, but are sequentially provided. An oil chamber 334 communicating with the outlet port 344 via the pilot oil passage 351 is provided between the auxiliary small-diameter portion 315 and the auxiliary small-diameter hole end surface 330. The first body 301a and the second body 301b are integrally assembled by a suitable method such as bolting (not shown) to form the body 301. At this time, the first body 301a side medium diameter hole 322 and the second body 301b Even if the side large-diameter hole 323 is misaligned, there is no operational problem since the second spool 312 and the third spool 310 are merely abutted as separate components.
[0043]
Then, the pressure compensating valve 41 applies the outlet pressure (Pz) of the outlet port 344 in the closing direction to the end face 340 (pressure receiving area B1) of the auxiliary small diameter portion 315 of the oil chamber 334 via the pilot oil passage 351 and the maximum load pressure port 346. Is applied to the pressure receiving area B2 of the oil chamber 336 obtained by subtracting the cross-sectional area of the auxiliary small-diameter portion 315 from the cross-sectional area of the second large-diameter land 314. Further, the pump discharge pressure (Pd) is applied to the pressure receiving area B1 of the first spool 311 of the oil chamber 331 through the auxiliary inlet port 341 in the opening direction of the pressure compensating valve 41, and the actuator load pressure (PL) of the load pressure port 342 to the It acts on the pressure receiving area B3 of the oil chamber 332 obtained by subtracting the sectional area B1 of the first spool 311 from the sectional area of the second spool 312. The cross-sectional area of the oil chamber 333, which is obtained by subtracting the cross-sectional area of the second spool 312 from the cross-sectional area of the first large-diameter land 313, communicates with the tank through the tank port 343. The working force does not work.
The sectional area B2 and the sectional area B1 of the first spool 311 are substantially equal (B1 = B2). In addition, the sectional area B3 is smaller than the sectional area B1 (= B2) of the first spool (B1> B3). ), The flow rate of the pressure compensating valve 41 connected to the actuator is reduced in accordance with the increase in the load pressure (PL) of each actuator.
[0044]
In operation, the force of the spring 350 pushing each spool is as weak as the spring 118 of FIG.
When each spool of the pressure compensating valve 41 is in an equilibrium state, the balance of the force applied to each spool is
Pz · B1 + Pm · B2 = Pd · B1 + PL · B3 (8)
Since B1 = B2,
Pz · B1 + Pm · B1 = Pd · B1 + PL · B3 (9)
Pz + Pm = Pd + PL · (B3 / B1)
If B1> B3 and (B3 / B1) = k, then
Pz + Pm = Pd + PL · k (10)
However, if k <1 k = 1− (1−k), then
Pz + Pm = Pd + PL. [1- (1-k)]
Pz + Pm = Pd + PL-PL · (1-k) (11)
Since ΔP = Pz−PL, from equation (11),
Pz-PL = -Pm + Pd-PL. (1-k)
ΔP = Pz−PL = (Pd−Pm) −PL · (1−k) (12)
Here, since Pc in FIG. 3 is Pc = Pd−Pm,
The pressure compensating valve 41 shown in FIG. 5 operates similarly to the pressure compensating valve 4 shown in FIG.
[0045]
For the same reason as described with reference to the pressure compensating valve 4 in FIG. 3, the value obtained by dividing the third pressure receiving area B3 of the two pressure compensating valves 41 and 42 by the first pressure receiving area B1 is 0.99-0. 95 (99-95%) is desirable. Also in the pressure compensating valves 41 and 42 of the hydraulic drive device according to the first invention of the present invention, at least two of the plurality of actuators are driven, such as driving two traveling motors that rotationally drive a pair of crawlers of a hydraulic traveling vehicle. When the two actuators need to be driven synchronously irrespective of the load pressure of the actuators, the third pressure receiving areas B3 of the two pressure compensating valves 41 and 42 communicating with the two actuators are changed to the first pressure receiving area B3. It is desirable to make the values divided by the area B1 equal.
Preferably, if the decrease rate of the flow rate is increased, it is likely to be overcorrected when straight ahead and meander easily, and it is difficult to operate while trying to maintain straightness when turning, and conversely if the decrease rate of the flow rate is small. Since the correction cannot be performed and the straightness cannot be secured, the value obtained by dividing the third pressure receiving area B3 of the two pressure compensating valves 41 and 42 by the first pressure receiving area B1 is 0.99 to 0.95. (99 to 95%). For the same reason as described for the pressure compensating valve 4 in FIG. 3, in the pressure compensating valves 41 and 42 used in the hydraulic drive device of the first invention of the present invention, at least two of the actuators 10 and 20 are used. The third pressure receiving area of the high load side pressure compensating valve 42 which communicates with the high load side actuator 20 when the load pressure of the high load side actuator 20 of the actuators is extremely larger than the load pressure of the other low load side actuator 10. A value obtained by dividing B3 by the first pressure receiving area B1 is made smaller than a value obtained by dividing the third pressure receiving area B3 of the low load side pressure compensating valve 41 communicating with the low load side actuator 10 by the first pressure receiving area B1. It is desirable. The value obtained by dividing the third pressure receiving area of the pressure compensating valve of the low load side actuator by the first pressure receiving area is 1 to 0.98, and the third pressure receiving area of the pressure compensating valve of the high load side actuator is the second pressure receiving area. It is desirable that the value divided by the pressure receiving area of 1 is 0.97 to 0.94.
[0046]
In the above embodiments, the hydraulic circuit that drives the two hydraulic actuators has been described. For example, in a hydraulic shovel, two traveling motors that rotationally drive a pair of crawlers of a hydraulic traveling vehicle, a turning hydraulic motor, and a boom, At least six actuators are operated, such as arm and bucket hydraulic cylinders. Therefore, in each of the embodiments described above, only two actuators representing these are shown. In the present invention, a plurality of hydraulic actuators are referred to as a plurality of traveling motors, a turning hydraulic motor, a hydraulic cylinder, and the like, respectively. And a plurality of pressure compensating and directional control valves, each communicating with a hydraulic actuator.
[Brief description of the drawings]
FIG. 1 is a hydraulic circuit diagram of a hydraulic drive device according to a first embodiment of the present invention.
FIG. 2 (a) is a view of the present invention.Of the second inventionFIG. 2 shows a hydraulic circuit diagram of a hydraulic drive device according to an embodiment, (b) is a partial hydraulic circuit diagram showing a pump discharge amount control circuit according to an embodiment different from (a), and (c) is different from (a). In the partial hydraulic circuit diagram showing the secondary pressure generating device, the other portions of FIG.
3 shows a sectional structural view (conceptual diagram) of an embodiment of a pressure compensating valve used in the hydraulic drive device of FIGS. 2 and 7. FIG.
FIG. 4 is a sectional structural view (conceptual diagram) of a pressure compensating valve of an embodiment different from FIG. 3 used in the hydraulic drive device of FIGS. 2 and 7;
FIG. 5 is a sectional structural view (conceptual diagram) of a pressure compensating valve used in the hydraulic drive device of FIG. 1;
FIG. 6 shows a hydraulic circuit diagram in which a pressure compensating valve is arranged downstream of a directional control valve according to an embodiment different from FIG. 2 (a).
FIG. 7A is a partial hydraulic circuit diagram of a hydraulic drive device that drives two traveling motors in synchronization with each other according to another embodiment different from FIG. 2A, and FIG. And FIG. 7 shows a partial hydraulic circuit diagram of a hydraulic drive device for driving two actuators having greatly different loads according to still another embodiment.
[Explanation of symbols]
1 prime mover 2 variable pump
4, 4 ', 14, 28, 29, 30, 36, 44, 48 Pressure compensating valve
8, 18 directional control valve 10, 11, 15, 15 ', 20, 25, 50, 51, hydraulic actuator
13 Shuttle valve 16 Maximum load pressure line
17 Pump volume changing means
19, 46 Spring 31 Differential pressure control valve
33, 35 Pilot oil passage 38, 45 Pump flow control valve

Claims

可変ポンプと、該可変ポンプの吐出油によって駆動される複数の油圧アクチュエータと、該複数のアクチュエータのそれぞれに流入する圧油を制御可能にされた流量調節機能を有する複数の方向制御弁及び各方向制御弁のそれぞれの圧力補償をする複数の圧力補償弁と、前記可変ポンプ吐出圧力（Ｐｄ）とアクチュエータの最高負荷圧力（Ｐｍ）との差圧に等しい二次圧力（Ｐｃ＝Ｐｄ−Ｐｍ）を発生する差圧制御弁と、該可変ポンプの吐出油を該可変ポンプの押しのけ容積変更手段に連通させるポンプ流量調整弁と、を有し、各前記圧力補償弁は圧力補償弁を閉じる方向に圧力補償弁の下流側の圧力（Ｐｚ）を作用させ、圧力補償弁を開く方向に該差圧制御弁から出力される二次圧力（Ｐｃ）及び方向制御弁の下流側圧力であるアクチュエータ負荷圧力（ＰＬ）を圧力補償弁を開く方向にそれぞれ作用させて前記圧力補償をするようにした油圧駆動装置において、該ポンプ流量調整弁のスプリングの作用力を該ポンプ流量調整弁を閉じ該可変ポンプの押しのけ容積を増大させる方向に作用させ、該二次圧力（Ｐｃ）を油路を介して該ポンプ流量調整弁を開き該可変ポンプの押しのけ容積を減少させるよう作用させるようにしたことを特徴とする油圧駆動装置。A variable pump, a plurality of hydraulic actuators driven by oil discharged from the variable pump, a plurality of directional control valves having a flow control function capable of controlling pressure oil flowing into each of the plurality of actuators, and each direction A plurality of pressure compensating valves for compensating the respective pressures of the control valves, and a secondary pressure (Pc = Pd-Pm) equal to the differential pressure between the variable pump discharge pressure (Pd) and the maximum load pressure (Pm) of the actuator. A differential pressure control valve to be generated, and a pump flow regulating valve for communicating the discharge oil of the variable pump to a displacement volume changing means of the variable pump. Each of the pressure compensating valves has a pressure in a direction to close the pressure compensating valve. Actuating the secondary pressure (Pc) output from the differential pressure control valve in the direction of opening the pressure compensating valve and the downstream pressure of the directional control valve by applying the pressure (Pz) downstream of the compensating valve. In the hydraulic drive device in which the load compensation pressure (PL) is applied in the direction in which the pressure compensating valve is opened to perform the pressure compensation, the working force of the spring of the pump flow regulating valve is reduced by closing the pump flow regulating valve. Acting in a direction to increase the displacement of the variable pump, and operating the secondary pressure (Pc) to decrease the displacement of the variable pump by opening the pump flow regulating valve via an oil passage. Features a hydraulic drive.

請求項１記載の油圧駆動装置において、前記圧力補償弁の少なくとも１は、そのアクチュエータ自身の自己の負荷圧力の増加に応じてそのアクチュエータに通じる圧力補償弁の流量を減少するようにしたことを特徴とする油圧駆動装置。2. The hydraulic drive device according to claim 1, wherein at least one of the pressure compensating valves is adapted to reduce a flow rate of the pressure compensating valve connected to the actuator according to an increase in a load pressure of the actuator itself. Hydraulic drive.

請求項２記載の油圧駆動装置において、各前記圧力補償弁の少なくとも１は対応する各前記方向制御弁の上流側に設けられ、該圧力補償弁は第１の油室の第１の受圧面積に自身の下流側の出口圧力を弁を閉じる方向に作用させ、第２の油室の第２の受圧面積に前記二次圧力を弁を開く方向に作用させ、そして第３の油室の第３の受圧面積に各前記アクチュエータの負荷圧力を弁を開く方向に作用させ、そして、前記第２の受圧面積と前記第３の受圧面積とをほぼ同じとし、かつ前記第１の受圧面積を前記第３の受圧面積より大きくしたことを特徴とする油圧駆動装置。3. The hydraulic drive device according to claim 2, wherein at least one of the pressure compensating valves is provided on the upstream side of each of the corresponding direction control valves, and the pressure compensating valve is provided on a first pressure receiving area of the first oil chamber. The second outlet pressure acts on the second pressure receiving area of the second oil chamber in the direction of opening the valve, and the third outlet of the third oil chamber acts on the second pressure receiving area of the second oil chamber. The load pressure of each of the actuators acts on the pressure receiving area in the direction in which the valve is opened, and the second pressure receiving area and the third pressure receiving area are substantially equal to each other, and the first pressure receiving area is the first pressure receiving area. 3. A hydraulic drive device characterized in that it is larger than the pressure receiving area of No. 3.

請求項３記載の油圧駆動装置において、前記圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値が０．９９〜０．９５（９９〜９５％）であることを特徴とする油圧駆動装置。4. The hydraulic drive device according to claim 3 , wherein a value obtained by dividing the third pressure receiving area of the pressure compensating valve by the first pressure receiving area is 0.99 to 0.95 (99 to 95%). Features a hydraulic drive.

請求項３記載の油圧駆動装置において、油圧走行車両の一対のクローラを回転駆動させる２個の走行モータでは、前記複数のアクチュエータのうちの少なくとも２個のアクチュエータがアクチュエータの負荷圧力に拘わらず同期して駆動する必要があるため、前記２個のアクチュエータに通ずる２個の前記圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値をそれぞれ等しくしたことを特徴とする油圧駆動装置。4. The hydraulic drive device according to claim 3 , wherein in the two traveling motors that rotationally drive a pair of crawlers of the hydraulic traveling vehicle, at least two of the plurality of actuators are synchronized regardless of the load pressure of the actuator. A hydraulic pressure, wherein values obtained by dividing the third pressure receiving area of the two pressure compensating valves connected to the two actuators by the first pressure receiving area are equal to each other. Drive.

請求項３記載の油圧駆動装置において、前記複数の油圧アクチュエータのうちの少なくとも２個のアクチュエータのうちの第１のアクチュエータの負荷圧力が他方の第２のアクチュエータの負荷圧力より極めて大きい負荷特性の場合、高負荷側のアクチュエータに通ずる高負荷特性側圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値を、低負荷側アクチュエータに通ずる低負荷側圧力補償弁の前記第３の受圧面積を前記第１の受圧面積で除した値より、大きくしたことを特徴とする油圧駆動装置。4. The hydraulic drive device according to claim 3 , wherein the load pressure of a first actuator of at least two of the plurality of hydraulic actuators has a load characteristic that is significantly higher than the load pressure of the other second actuator. A value obtained by dividing the third pressure receiving area of the high load characteristic side pressure compensating valve communicating with the high load side actuator by the first pressure receiving area, and calculating the value of the low load side pressure compensating valve communicating with the low load side actuator. 3. The hydraulic drive device according to claim 3, wherein the pressure receiving area is larger than a value obtained by dividing the pressure receiving area by the first pressure receiving area.

請求項６記載の油圧駆動装置において、前記低負荷側アクチュエータの圧力補償弁の第３の受圧面積を第１の受圧面積で除した値が１〜０．９８、前記高負荷側アクチュエータの圧力補償弁の第３の受圧面積を第１の受圧面積で除した値が０．９７〜０．９４であることを特徴とする油圧駆動装置。 7. The hydraulic drive device according to claim 6 , wherein a value obtained by dividing a third pressure receiving area of the pressure compensating valve of the low load side actuator by a first pressure receiving area is 1 to 0.98, and the pressure compensation of the high load side actuator is performed. A hydraulic drive device wherein a value obtained by dividing a third pressure receiving area of the valve by the first pressure receiving area is 0.97 to 0.94.

請求項３記載の油圧駆動装置において、前記圧力補償弁は、本体と、該本体に設けた小径本体穴及びそれに続く大径本体穴の２つの内径を有する本体穴に摺動可能にそれぞれ嵌合する小径部及び第１及び第２のの大径ランドを有するスプールと、該本体穴に沿って該本体に順次設けれれたアクチュエータの負荷圧力ポート、二次圧力ポート、出口ポート、ポンプ吐出油路と連通する入口ポート及びタンクポートとを有し、該小径本体穴に嵌合する該スプールの一端に設けた前記小径部はスプリングを介して本体穴端面に当接可能にかつ該本体穴との間で該負荷圧力ポートに通じる該第３の油室を形成しそして該スプールの他端は該タンクポートに通じるタンク油室を形成し、該スプールの該小径部と該第１の大径ランドとの接合部を囲む該大径本体穴内に該二次圧力ポートに通じる該第２の油室を形成し、該スプールの他端に設けられた軸方向穴にピストンが油密に入れ子式に摺動可能に挿入されかつ該ピストンの他端はもう一方の本体穴端面に当接可能にされて該タンク油室内に位置し、該軸方向穴内の該スプールと該ピストンとの間にパイロット油路を介して該出口ポートに通じる該第１の油室を形成しており、該第１の油室の第１の受圧面積Ａ１は該ピストンの断面積により、該第２の油室の第２の受圧面積Ａ２は該大径本体穴断面積から該小径穴断面積を引いた面積により、そして該第３の油室の第３の受圧面積Ａ３は該小径部断面積により、それぞれ形成させ、かつ該スプールには、該第１の大径ランド部に面する該第２の大径ランドに設けた該出口ポートと入口ポートとの間を絞る開閉可能な絞り部と、を有し、
前記第２の受圧面積Ａ２と前記第３の受圧面積Ａ３とをほぼ同じにし、かつ前記第３の受圧面積Ａ３を前記第１の受圧面積Ａ１より小にし、そのアクチュエータ自身の自己の負荷圧力（ＰＬ）の増加に応じてそのアクチュエータに通じる前記圧力補償弁の流量を減少するようにしたことを特徴とする油圧駆動装置。4. The hydraulic drive device according to claim 3 , wherein the pressure compensating valve is slidably fitted into a main body, and a main body hole having two inner diameters of a small-diameter main body hole provided in the main body and a large-diameter main body hole. A spool having a small-diameter portion and first and second large-diameter lands, and a load pressure port, a secondary pressure port, an outlet port, and a pump discharge oil passage of an actuator sequentially provided in the main body along the main body hole. The small diameter portion provided at one end of the spool that fits into the small diameter main body hole can be in contact with an end face of the main body hole via a spring, and is connected to the main body hole. And the other end of the spool forms a tank oil chamber that communicates with the tank port, the small diameter portion of the spool and the first large diameter land. The large area surrounding the junction with The second oil chamber communicating with the secondary pressure port is formed in the body hole, and a piston is oil-tightly nested and slidably inserted into an axial hole provided at the other end of the spool. Is located in the tank oil chamber so as to be in contact with the end face of the other body hole, and communicates with the outlet port through a pilot oil passage between the spool and the piston in the axial hole. The first oil chamber is formed, and the first pressure receiving area A1 of the first oil chamber is determined by the cross-sectional area of the piston, and the second pressure receiving area A2 of the second oil chamber is determined by the large diameter. The third pressure receiving area A3 of the third oil chamber is formed by the area obtained by subtracting the small-diameter hole cross-sectional area from the body hole cross-sectional area, and the third oil-receiving chamber is formed by the small-diameter portion cross-sectional area. The outlet port and the inlet port provided on the second large-diameter land facing the first large-diameter land portion; Anda openable throttle portion for throttling between,
The second pressure receiving area A2 and the third pressure receiving area A3 are made substantially the same, and the third pressure receiving area A3 is made smaller than the first pressure receiving area A1. A hydraulic drive device wherein the flow rate of the pressure compensating valve communicating with the actuator is reduced in accordance with an increase in PL).

請求項３記載の油圧駆動装置において、前記圧力補償弁は、本体と、該本体に設けた本体穴と、該本体穴に摺動可能に嵌合された第１、第２及び第３の大径ランドを有するスプールと、該本体穴に沿って該本体に順次設けられた二次圧力ポート、アクチュエータの負荷圧力ポート、出口ポート、ポンプ吐出油路と連通する入口ポート、及びタンクポートとを有し、前記スプールの一端に設けられた補助軸方向穴に補助ピストンが油密に入れ子式に摺動可能に挿入されかつ該補助ピストンの他端は該本体穴の端面に当接可能にされて両者間に該二次圧力ポートに通じる該第２油室を形成し、該補助軸方向穴内の前記スプールと該補助ピストンとの間にスプリングが介されかつ補助パイロット油路を介して該負荷圧力ポートに通じる該第３油室を形成し、そして前記スプールの他端は該本体穴の他端面との間にタンクポートに通じるタンク油室を形成し、前記スプールの他端に設けられた軸方向穴にピストンが油密に入れ子式に摺動可能に挿入されかつ該ピストンの他端は該本体穴のもう一方の端面に当接可能にされ該タンク油室内に位置し、該軸方向穴内の前記スプールと該ピストンとの間にパイロット油路を介して該出口圧力ポートに通じる該第１油室を形成し、前記第１油室の第１受圧面積は該ピストンの断面積により、前記第２油室の第２受圧面積は該大径穴断面積から該補助ピストン断面積を引いた面積により、そして前記第３油室の第３受圧面積は該補助ピストン断面積により、それぞれ形成させ、かつ、前記スプールには、該第２の大径ランド部に面する前記第３の大径ランドに設けた前記出口ポートと入口ポート間を絞る開閉可能な絞り部と、を有し、前記第２受圧面積と前記第３受圧面積とをほぼ同じにし、かつ前記第３の受圧面積を前記第１の受圧面積Ａ１より小にしそのアクチュエータ自身の自己の負荷圧力の増加に応じてそのアクチュエータに通じる前記圧力補償弁の流量を減少するようにしたことを特徴とする油圧駆動装置。4. The hydraulic drive device according to claim 3 , wherein the pressure compensating valve includes a main body, a main body hole provided in the main body, and first, second and third large-sized slidably fitted in the main body hole. A spool having a diameter land, a secondary pressure port, a load pressure port of the actuator, an outlet port, an inlet port communicating with the pump discharge oil passage, and a tank port sequentially provided in the main body along the main body hole. An auxiliary piston is oil-tightly and slidably inserted into an auxiliary axial hole provided at one end of the spool, and the other end of the auxiliary piston is made to be able to contact an end face of the main body hole. The second oil chamber communicating with the secondary pressure port is formed therebetween, and a spring is provided between the spool and the auxiliary piston in the auxiliary axial hole and the load pressure is set via an auxiliary pilot oil path. The third oil chamber leading to the port The other end of the spool forms a tank oil chamber communicating with the tank port between the other end of the main body hole, and a piston is oil-tightly nested in an axial hole provided at the other end of the spool. The piston is slidably inserted and the other end of the piston is allowed to abut the other end face of the body hole and is located in the tank oil chamber, between the spool and the piston in the axial hole. A first oil chamber communicating with the outlet pressure port via a pilot oil passage, wherein a first pressure receiving area of the first oil chamber is a second pressure receiving area of the second oil chamber depending on a cross-sectional area of the piston. Is formed by the area obtained by subtracting the auxiliary piston cross-sectional area from the large-diameter hole cross-sectional area, and the third pressure receiving area of the third oil chamber is formed by the auxiliary piston cross-sectional area. The third large diameter facing the second large diameter land portion; And an openable and closable throttle portion for narrowing a gap between the outlet port and the inlet port provided in the cylinder, making the second pressure receiving area substantially equal to the third pressure receiving area, and setting the third pressure receiving area to A hydraulic drive device wherein the flow rate of the pressure compensating valve, which is smaller than the first pressure receiving area A1 and communicates with the actuator according to the increase of its own load pressure, is reduced.

可変ポンプと、該可変ポンプの吐出油によって駆動される複数の油圧アクチュエータと、該複数のアクチュエータのそれぞれに流入する圧油を制御可能にされた流量調節機能を有する複数の方向制御弁及び各方向制御弁のそれぞれの圧力補償をする複数の圧力補償弁とを有し、各前記圧力補償弁は閉じ方向に圧力補償弁の下流側の圧力（Ｐｚ）及び前記複数のアクチュエータのうちの最高負荷圧力（Ｐｍ）をそれぞれ作用させ、圧力補償弁の開方向に圧力補償弁の上流側の圧力であるポンプ吐出圧力（Ｐｄ）及び方向制御弁の下流側圧力であるアクチュエータ負荷圧力（ＰＬ）をそれぞれ作用させて前記圧力補償をし、該可変ポンプの吐出油を該可変ポンプの押しのけ容積変更手段に連通させるポンプ流量調整弁を有し、該最高負荷圧力（Ｐｍ）を油路を介して該ポンプ流量調整弁を閉じ該可変ポンプの押しのけ容積を増大させる方向に該ポンプ流量調整弁のスプリングの作用力とともに作用させ、該ポンプ吐出圧力（Ｐｄ）を別の油路を介して該ポンプ流量調整弁を開き該可変ポンプの押しのけ容積を減少させ方向に作用させるようにされた油圧駆動装置において、
前記圧力補償弁の少なくとも１は、相互に連結された第１ボディと第２ボディとからなるボデイと、該第１ボデイに設けた小径穴と、小径穴に続く中径穴と、該小径穴に嵌合する第１スプールと、該中径穴に嵌合する第２スプールと、該第２ボデイに設けた該中径穴に隣接した大径穴及び大径穴に続く該小径穴と同径の補助小径穴と、該大径穴に嵌合する第１及び第２の大径ランド及び該補助小径穴に嵌合する補助小径部を有する第３スプールと、を有し、該第１スプールと該ボデイ小径穴端面間に前記各スプールを押すスプリングを介し、さらに、該ボデイに沿って順次設けた、ポンプ吐出油路と連通し該小径穴に通じる補助入口ポート、アクチュエータの負荷圧力ラインに連通しかつ該中径穴に通ずるアクチュエータの負荷圧力ポート、該第２スプールと第３スプールとの当接部で該大径穴に通じるタンクポート、該第１及び第２の大径ランド間の該大径穴に通じる出口ポート、該ポンプ吐出油路と連通しかつ該第２の大径ランドに設けた開閉可能な絞り部で絞られるように開口する入口ポート、及び該アクチュエータのうちの最高負荷圧力取出ラインと連通しかつ該第２の大径ランドと該補助小径部との連続部の該大径穴に通ずる最高負荷圧力ポート、とを有し、該補助小径部と該補助小径穴端面間にパイロット油路を介して該出口ポートと連通する油室を設け、そして、前記圧力補償弁は閉じ方向に該出口ポート圧力（Ｐｚ）をパイロット油路を介して該補助小径部の端面（断面積Ｂ１）に、該最高負荷圧力ポートの最高負荷圧力（Ｐｍ）を該第２の大径ランドの断面積から該補助小径部の断面積を差し引いた断面積Ｂ２に、そして逆に該圧力補償弁を開く方向に該補助入口ポートを介して該ポンプ吐出圧力（Ｐｄ）を該第１スプールの断面積Ｂ１に、該負荷圧力ポートのアクチュエータ負荷圧力（ＰＬ）を該第２スプールの断面積から該第１スプールの断面積Ｂ１を差し引いた断面積Ｂ３に、それぞれ作用させ、さらに、前記断面積Ｂ２と前記第１スプールの断面積Ｂ１とをほぼ同じとし（Ｂ１＝Ｂ２）、前記断面積Ｂ３を前記第１スプールの断面積Ｂ１（＝Ｂ２）より小にし（Ｂ１＞Ｂ３）、そのアクチュエータ自身の自己の負荷圧力（ＰＬ）の増加に応じてそのアクチュエータに通じる前記圧力補償弁の流量を減少するようにしたことを特徴とする油圧駆動装置。A variable pump, a plurality of hydraulic actuators driven by oil discharged from the variable pump, a plurality of directional control valves having a flow control function capable of controlling pressure oil flowing into each of the plurality of actuators, and each direction A plurality of pressure compensating valves for compensating the respective pressures of the control valves, wherein each of the pressure compensating valves is a pressure (Pz) downstream of the pressure compensating valve in a closing direction and a maximum load pressure of the plurality of actuators. (Pm), and the pump discharge pressure (Pd), which is the pressure upstream of the pressure compensating valve, and the actuator load pressure (PL), which is the downstream pressure of the directional control valve, act in the opening direction of the pressure compensating valve. And a pump flow regulating valve for communicating the discharge oil of the variable pump with the displacement volume changing means of the variable pump. ) Is acted together with the force of the spring of the pump flow control valve in a direction to close the pump flow control valve via the oil passage in a direction to increase the displacement of the variable pump, and to change the pump discharge pressure (Pd) to another oil. A hydraulic drive device adapted to open the pump flow control valve through a passage to act in a direction to reduce the displacement of the variable pump;
At least one of the pressure compensating valves includes a body including a first body and a second body connected to each other, a small diameter hole provided in the first body, a medium diameter hole following the small diameter hole, and a small diameter hole. A first spool fitted in the second bore, a second spool fitted in the middle bore, and a large bore adjacent to the middle bore provided in the second body and a small bore following the large bore. An auxiliary small-diameter hole having a diameter, a third spool having first and second large-diameter lands fitted into the large-diameter hole, and an auxiliary small-diameter portion fitted into the auxiliary small-diameter hole; An auxiliary inlet port which is provided between the spool and the end face of the body small-diameter hole via a spring for pushing the spool, and which is sequentially provided along the body and communicates with the pump discharge oil passage and communicates with the small-diameter hole; Load pressure port of the actuator communicating with the medium diameter hole and A tank port communicating with the large-diameter hole at a contact portion between the second spool and the third spool, an outlet port communicating between the first and second large-diameter lands with the large-diameter hole, and communicating with the pump discharge oil passage; And an inlet port which is opened so as to be throttled by an openable and closable throttle provided on the second large-diameter land, and which communicates with a maximum load pressure extraction line of the actuator and which is connected to the second large-diameter land. An oil port having a maximum load pressure port communicating with the large-diameter hole in a continuous portion with the auxiliary small-diameter portion, and communicating with the outlet port via a pilot oil passage between the auxiliary small-diameter portion and an end face of the auxiliary small-diameter hole. A pressure chamber, and the pressure compensating valve applies the outlet port pressure (Pz) in a closing direction to an end face (cross-sectional area B1) of the auxiliary small diameter portion via a pilot oil passage, to a maximum load pressure of the maximum load pressure port. (Pm) is the sectional area of the second large-diameter land. The pump discharge pressure (Pd) is reduced to a cross-sectional area B1 of the first spool through the auxiliary inlet port in a direction to open the pressure compensating valve to a cross-sectional area B2 obtained by subtracting a cross-sectional area of the auxiliary small diameter portion. And the actuator load pressure (PL) of the load pressure port is applied to a cross-sectional area B3 obtained by subtracting a cross-sectional area B1 of the first spool from a cross-sectional area of the second spool. The cross-sectional area B1 of one spool is made substantially the same (B1 = B2), the cross-sectional area B3 is made smaller than the cross-sectional area B1 (= B2) of the first spool (B1> B3), and the actuator itself loads itself. A hydraulic drive device wherein the flow rate of the pressure compensating valve connected to the actuator is reduced in accordance with an increase in the pressure (PL).

請求項１０記載の油圧駆動装置において、前記圧力補償弁の前記断面積Ｂ３を前記断面積Ｂ１で除した値が０．９９〜０．９５（９９〜９５％）であることを特徴とする油圧駆動装置。The hydraulic drive according to claim 10 , wherein a value obtained by dividing the sectional area B3 of the pressure compensating valve by the sectional area B1 is 0.99 to 0.95 (99 to 95%). Drive.

請求項１０記載の油圧駆動装置において、油圧走行車両の一対のクローラを回転駆動させる２個の走行モータでは、前記複数のアクチュエータのうちの少なくとも２個のアクチュエータがアクチュエータの負荷圧力に拘わらず同期して駆動する必要があるため、前記２個のアクチュエータに通ずる２個の前記圧力補償弁の前記断面積Ｂ３を前記断面積Ｂ１で除した値をそれぞれ等しくしたことを特徴とする油圧駆動装置。11. The hydraulic drive device according to claim 10 , wherein in the two traveling motors that rotationally drive a pair of crawlers of the hydraulic traveling vehicle, at least two of the plurality of actuators are synchronized regardless of the load pressure of the actuator. A hydraulic drive device wherein the values obtained by dividing the cross-sectional areas B3 of the two pressure compensating valves communicating with the two actuators by the cross-sectional area B1 are equal to each other.

請求項１０記載の油圧駆動装置において、前記複数の油圧アクチュエータのうちの少なくとも２個のアクチュエータのうちの第１のアクチュエータの負荷圧力が他方の第２のアクチュエータの負荷圧力より極めて大きい負荷特性の場合、高負荷側のアクチュエータに通ずる高負荷特性側圧力補償弁の前記断面積Ｂ３を前記断面積Ｂ１で除した値を、低負荷側アクチュエータに通ずる低負荷側圧力補償弁の前記断面積Ｂ３を前記断面積Ｂ１で除した値より、小さくしたことを特徴とする油圧駆動装置。The hydraulic drive device according to claim 10 , wherein a load characteristic of a first actuator of at least two of the plurality of hydraulic actuators has a load characteristic which is significantly higher than a load pressure of the other second actuator. The value obtained by dividing the cross-sectional area B3 of the high-load characteristic-side pressure compensating valve communicating with the high-load-side actuator by the cross-sectional area B1 is referred to as the low-load-side pressure compensating valve communicating with the low-load-side actuator. A hydraulic drive device characterized in that the value is smaller than a value obtained by dividing the sectional area by B1.

請求項１３記載の油圧駆動装置において、前記低負荷側アクチュエータの圧力補償弁の前記断面積Ｂ３を前記断面積Ｂ１で除した値が１〜０．９８、前記高負荷側アクチュエータの圧力補償弁の前記断面積Ｂ３を前記断面積Ｂ１で除した値が０．９７〜０．９４であることを特徴とする油圧駆動装置。14. The hydraulic drive device according to claim 13 , wherein a value obtained by dividing the sectional area B3 of the pressure compensating valve of the low load side actuator by the sectional area B1 is 1 to 0.98, and that of the pressure compensating valve of the high load side actuator. A hydraulic drive device wherein a value obtained by dividing the sectional area B3 by the sectional area B1 is 0.97 to 0.94.