JP3907040B2

JP3907040B2 - Hydraulic drive device for hydraulic excavator

Info

Publication number: JP3907040B2
Application number: JP2001304844A
Authority: JP
Inventors: 靖貴釣賀; 純也川本; 究高橋
Original assignee: Nachi Fujikoshi Corp; Hitachi Construction Machinery Co Ltd
Current assignee: Nachi Fujikoshi Corp; Hitachi Construction Machinery Co Ltd
Priority date: 2001-10-01
Filing date: 2001-10-01
Publication date: 2007-04-18
Anticipated expiration: 2021-10-01
Also published as: JP2003113804A

Description

【０００１】
【発明の属する技術分野】
本発明は、油圧ショベルに用いられるＬＳ制御方式の油圧駆動装置に係わり、特に、油圧ポンプの吐出圧が複数のアクチュエータの最高負荷圧より目標差圧だけ高くなるようロードセンシング制御しかつ複数の流量制御弁の前後差圧を制御する圧力補償弁のそれぞれの目標補償差圧を油圧ポンプの吐出圧と複数のアクチュエータの最高負荷圧との差圧により設定すると共に、ロードセンシング制御の目標差圧をエンジン回転数に依存する可変値として設定する油圧ショベルの油圧駆動装置に関する。
【０００２】
【従来の技術】
油圧ポンプの吐出圧が複数のアクチュエータの最高負荷圧より目標差圧だけ高くなるようロードセンシング制御する油圧駆動装置はロードセンシングシステム（以下、適宜ＬＳシステムという）と呼ばれ、複数の流量制御弁の前後差圧をそれぞれ圧力補償弁により制御し、複数のアクチュエータを同時に駆動する複合操作時に負荷圧の大小に係わらず流量制御弁の開口面積に応じた比率で圧油を供給できるようにしている。
【０００３】
このようなＬＳシステムでは、油圧ポンプの吐出圧と複数のアクチュエータの最高負荷圧との差圧（以下、ＬＳ差圧という）を圧力補償弁に導き、圧力補償弁のそれぞれの目標補償差圧をＬＳ差圧により設定することが一般的に行われており、これにより複数のアクチュエータを同時に駆動する複合動作時に、油圧ポンプの吐出流量が複数の流量制御弁の要求する流量に満たないサチュレーション状態になったときでも、サチュレーションの程度に応じてＬＳ差圧が低下し、これに伴って圧力補償弁の目標補償差圧も小さくなるため、油圧ポンプの吐出流量をそれぞれのアクチュエータが要求する流量の比に再分配することができる。
【０００４】
このようなＬＳシステムにおいて、特開平１０−８９３０４号公報には、油圧ポンプの吐出圧と複数のアクチュエータの最高負荷圧との差圧を絶対圧として出力する差圧減圧弁を設け、この差圧減圧弁の出力圧を複数の圧力補償弁に導き、それぞれの目標補償差圧を設定するようにしたものが記載されている。
【０００５】
また、特開平１０−１９６６０４号公報には、油圧ポンプを駆動するエンジンの回転数に依存する圧力を絶対圧として出力する差圧減圧弁を設け、この差圧減圧弁の出力圧をＬＳ制御レギュレータに導き、ロードセンシング制御の目標差圧（以下、目標ＬＳ差圧という）をエンジンの回転数に依存する可変値として設定した油圧駆動装置が記載されている。
【０００６】
更に、特開２０００−２２７１０３号公報には、旋回セクションの圧力補償弁に、旋回負加圧が上昇すると目標補償差圧が小さくなる負荷依存特性を持たせ旋回モータのスムーズな加速及び定常状態への移行を可能とすると共に、目標補償差圧が規定値以下にならないようにする下限設定バネ（旋回優先バネ）を設け、旋回複合操作での起動時にポンプ吐出流量がサチュレーション状態となっても、他のアクチュエータに比べ旋回速度が極端に遅くならないようにした油圧駆動装置が記載されている。
【０００７】
【発明が解決しようとする課題】
しかしながら上記従来技術には、１つのアクチュエータを駆動する単独動作から複数のアクチュエータを同時に駆動する複合動作に移行する場合に、操作性の低下が発生するという問題がある。
【０００８】
つまり、複数のアクチュエータを同時に駆動する複合動作で、動作している全てのアクチュエータの要求流量の総和が油圧ポンプの吐出可能な最大流量を越えた場合、サチュレーション状態が発生する。この状態では、サチュレーションの程度に応じてＬＳ差圧が低下し、これに伴って圧力補償弁の目標補償差圧も小さくなるため、各流量制御弁では油圧ポンプの吐出流量をそれぞれのアクチュエータが要求する流量の比に再分配することができる。これにより各流量制御弁に対する操作レバーの入力比に応じて全体のアクチュエータスピードが低下することになり、複合動作時に、特定のアクチュエータが停止してしまうことなく、操作性を維持することが可能になる。
【０００９】
しかし、この複合動作において、全てのアクチュエータの速度が均等に低下するため、操作性が低下する場合がある。
【００１０】
例えば、油圧ショベルの旋回操作ではオペレータが速度変化を体感できるため、速度変化に敏感であり、他のアクチュエータに比べ速度変化を抑えた構成とする必要がある。また、吊り荷作業等でも、旋回中に速度変化が生じると、荷がオペレータの意図に反して揺れてしまう。
【００１１】
特開２０００−２２７１０３号公報のように、旋回優先バネを圧力補償弁に設定する方法も考えられる。しかし、この場合はエンジン回転数によって目標補償差圧が変化するため、全てのエンジン回転数において、同様の割合で優先性を維持することが不可能である。
【００１２】
本発明の目的は、ＬＳシステムを備えた油圧ショベルの油圧駆動装置において、ポンプ吐出流量のサチュレーション状態が生じても旋回モータに優先的に圧油を供給してその速度変化を抑えることができ、かつエンジン回転数の設定に係わらず同様に優先性を維持することができ、優れた操作性を実現できるようにしたものを提供することである。
【００１３】
【課題を解決するための手段】
（１）上記目的を達成するために、本発明は、エンジンと、このエンジンにより駆動される可変容量型の油圧ポンプと、この油圧ポンプから吐出された圧油により駆動される旋回モータを含む複数のアクチュエータと、前記油圧ポンプから複数のアクチュエータに供給される圧油の流量を制御する複数の流量制御弁と、前記複数の流量制御弁の前後差圧をそれぞれ制御する複数の圧力補償弁と、前記油圧ポンプの吐出圧が前記複数のアクチュエータの最高負荷圧より目標差圧だけ高くなるようロードセンシング制御するポンプ制御手段と、前記エンジンの回転数が低下するにしたがって低下するよう前記エンジンの回転数に依存する圧力を出力するエンジン回転数検出回路とを備え、このエンジン回転数検出回路の出力圧を前記ポンプ制御手段に導き、前記エンジン回転数検出回路の出力圧により前記ロードセンシング制御の目標差圧を設定することで、前記エンジンの回転数が低下するにしたがって前記ロードセンシング制御の目標差圧が低下するよう前記ロードセンシング制御の目標差圧を前記エンジンの回転数に依存する可変値として設定した油圧ショベルの油圧駆動装置において、前記旋回モータ以外のアクチュエータに係わる圧力補償弁のそれぞれの目標補償差圧を前記油圧ポンプの吐出圧と前記複数のアクチュエータの最高負荷圧との差圧により設定すると共に、前記旋回モータに係わる圧力補償弁に前記エンジン回転数検出回路の出力圧を導き、前記エンジン回転数検出回路の出力圧により目標補償差圧を設定し、前記ポンプ制御手段は、前記油圧ポンプの最少吐出流量が前記旋回モータに係わる流量制御弁の最大要求流量より小さくならないように前記油圧ポンプの傾転を制限するリミッタ手段を備えるポンプ傾転可変機構を有するものとする。
【００１４】
このように旋回モータ以外のアクチュエータに係わる圧力補償弁のそれぞれの目標補償差圧をＬＳ差圧により設定し、旋回モータに係わる圧力補償弁の目標補償差圧をエンジン回転数検出回路の出力圧（ロードセンシング制御の目標差圧）により設定することにより、旋回モータを含む複合動作時（サチュレーション状態）に、旋回モータ以外のアクチュエータでは、ＬＳ差圧の低下に応じて圧力補償弁の目標補償差圧が低下しアクチュエータスピードが低下するのに対して、旋回モータ側では圧力補償弁にサチュレーション情報が伝わらず、目標補償差圧はサチュレーション状態でないときと同じままであり、旋回モータに優先的に圧油が供給されそのスピードダウンを抑えることが可能になる。
【００１５】
また、エンジン回転数を下げた場合は、エンジン回転数検出回路の出力圧（ロードセンシング制御の目標差圧）がそれに応じて低下し、それに伴い油圧ポンプの吐出圧と最高負荷圧の差圧も低下するよう制御され、旋回モータ以外のアクチュエータに係わる圧力補償弁の目標補償差圧が低下するだけでなく、旋回モータに係わる圧力補償弁の目標補償差圧も同様に低下する。その結果、旋回モータが速くなりすぎるというようなことが起こらず、エンジン回転数の設定によらず、旋回の適切な優先性を維持することができる。
【００１６】
更に、ポンプ制御手段にポンプ傾転可変機構を設け、このポンプ傾転可変機構に、油圧ポンプの最少吐出流量が旋回モータに係わる流量制御弁の最大要求流量より小さくならないように油圧ポンプの傾転を制限するリミッタ手段を設けることにより、特定のアクチュエータの単独動作時に、アクチュエータ速度が目標速度に達し油圧ポンプの吐出流量が減るようにロードセンシング制御されるとき、油圧ポンプの吐出流量が最少吐出流量に達すると、それ以降は最少吐出流量に維持され、油圧ポンプのＬＳ制御は動作しなくなる。つまり、油圧ポンプは最少吐出流量を容量とする固定ポンプであるかの如く動作する。その結果、油圧ポンプのＬＳ制御と旋回圧力補償弁の干渉を回避することができ、システムの安定性を維持することが可能になる。
【００１８】
（２）また、上記（１）において、好ましくは、前記エンジン回転数検出回路は、前記エンジンの回転数に依存する圧力を絶対圧として出力する第１差圧減圧弁を有し、前記旋回モータ以外のアクチュエータに係わる圧力補償弁に前記第１差圧減圧弁の出力圧を導きそれぞれの目標補償差圧を設定する。
【００１９】
（３）更に、上記（１）において、好ましくは、前記油圧ポンプの吐出圧と前記複数のアクチュエータの最高負荷圧との差圧を絶対圧として出力する第２差圧減圧弁を更に備え、前記旋回モータに係わる圧力補償弁に前記第２差圧減圧弁の出力圧を導き、この第２差圧減圧弁の出力圧で当該圧力補償弁の目標補償差圧を設定する。
【００２０】
【発明の実施の形態】
以下、本発明の実施の形態を図面を用いて説明する。
【００２１】
図１は本発明の一実施の形態に係わる油圧駆動装置を示す図である。
【００２２】
図１において、本実施の形態に係わる油圧駆動装置は、エンジン１と、このエンジン１により駆動されるメインポンプとしての可変容量型の油圧ポンプ２及び固定容量型のパイロットポンプ３０と、メインの油圧ポンプ２から吐出された圧油により駆動される複数のアクチュエータ３ａ，３ｂ，３ｃと、油圧ポンプ２の供給油路５に接続され、油圧ポンプ２からアクチュエータ３ａ，３ｂ，３ｃに供給される圧油の流量と方向をそれぞれ制御する複数の流量制御弁（メインスプール）４ａ，４ｂ，４ｃ及び油圧ポンプ２の吐出圧と複数のアクチュエータ３ａ，３ｂ，３ｃの最高負荷圧との差圧（ＬＳ差圧）を絶対圧として出力する差圧減圧弁１１を含むコントロールバルブ４と、油圧ポンプ２の傾転（容量）を制御するポンプ傾転制御機構１２と、エンジン回転数に依存する圧力を絶対圧として出力する差圧減圧弁５１を含むエンジン回転数検出回路１３とを備えている。
【００２３】
本実施の形態に係わる油圧駆動装置は例えば油圧ショベルに搭載されるものであり、アクチュエータ３ａ，３ｂ，３ｃは例えばそれぞれ油圧ショベルの旋回モータ、ブームシリンダ、アームシリンダである。油圧ショベルは下部走行体に旋回可能に搭載された上部旋回体と、上部旋回体に上下方向に回動可能に装備されたブーム、アーム、バケットからなるフロント作業機構を有し、旋回モータ３ａは上部旋回体を旋回駆動し、ブームシリンダ３ｂ及びアームシリンダ３ｃはそれぞれブーム及びアームを上下方向に回動駆動する。
【００２４】
複数の流量制御弁４ａ，４ｂ，４ｃは、それぞれ、クローズドセンタ型の複数の流量制御弁６ａ，６ｂ，６ｃと、これら複数の流量制御弁６ａ，６ｂ，６ｃのメータイン絞り部６１，６２の前後差圧を同じ値に制御する複数の圧力補償弁７ａ，７ｂ，７ｃとで構成されている。
【００２５】
流量制御弁６ａ，６ｂ，６ｃはそれぞれ図示しない操作レバーの操作により切り換え操作され、その操作レバーの操作量に応じてメータイン絞り部６１又は６２の開口面積が決まる。
【００２６】
複数の圧力補償弁７ａ，７ｂ，７ｃは、それぞれ、流量制御弁６ａ，６ｂ，６ｃのメータイン絞り部６１，６２の上流に設置された前置きタイプ（ビフォアオリフィスタイプ）であり、圧力補償弁７ａは１対の対向する受圧部７０ａ，７０ｂと開方向作動の受圧部７０ｃとを有し、受圧部７０ａ，７０ｂに流量制御弁６ａの上流側及び下流側の圧力がそれぞれ導かれ、受圧部７０ｃに差圧減圧弁５１の出力圧が導かれ、その出力圧を目標補償差圧として流量制御弁６ａの前後差圧を制御する。圧力補償弁７ｂは、１対の対向する受圧部７１ａ，７１ｂと開方向作動の受圧部７１ｃとを有し、受圧部７１ａ，７１ｂに流量制御弁６ｂの上流側及び下流側の圧力がそれぞれ導かれ、受圧部７１ｃに差圧減圧弁１１の出力圧が導かれ、その出力圧を目標補償差圧として流量制御弁６ｂの前後差圧を制御する。圧力補償弁７ｃも圧力補償弁７ｂと同様であり、１対の対向する受圧部７２ａ，７２ｂと開方向作動の受圧部７２ｃとを有し、受圧部７２ａ，７２ｂに流量制御弁６ｃの上流側及び下流側の圧力がそれぞれ導かれ、受圧部７２ｃに差圧減圧弁１１の出力圧が導かれ、その出力圧を目標補償差圧として流量制御弁６ｃの前後差圧を制御する。
複数の流量制御弁６ａ，６ｂ，６ｃには、それぞれ、アクチュエータ３ａ，３ｂ，３ｃの駆動時にそれらの負荷圧を取り出す負荷ポート６０ａ，６０ｂ，…が設けられ、これら負荷ポート６０ａ，６０ｂ，…に取り出された負荷圧のうちの最高の圧力が負荷ライン８ａ，８ｂ，８ｃ、８ｄ及びシャトル弁９ａ，９ｂを介して信号ライン１０に検出される。
【００２７】
差圧減圧弁１１は、出力圧を増やす側に位置する受圧部１１ａと出力圧を減らす側に位置する受圧部１１ｂ，１１ｃを有し、受圧部１１ａに油圧ポンプ２の吐出圧が導かれ、受圧部１１ｂ，１１ｃにそれぞれ信号ライン１０に検出された最高負荷圧と自己の出力圧が導かれ、これらの圧力のバランスで油圧ポンプ２の吐出圧と最高負荷圧との差圧（ＬＳ差圧）を絶対圧として出力する。
【００２８】
差圧減圧弁１１の出力ポートは信号ライン２１，２２を介してポンプ傾転制御機構１２に設けられたＬＳ制御弁１２ｂの受圧部１２ｄに接続され、差圧減圧弁１１の出力圧が受圧部１２ｄに導かれる。また、差圧減圧弁１１の出力ポートは、信号ライン２１，２４を介して圧力補償弁７ｂの受圧部７１ｃに接続され、信号ライン２１，２５を介して圧力補償弁７ｃの受圧部７２ｃに接続され、差圧減圧弁１１の出力圧が上記のように目標補償差圧として受圧部７１ｃ，７２ｃに導かれる。
【００２９】
ポンプ傾転制御機構１２は、油圧ポンプ２の吐出圧が高くなると油圧ポンプ２の傾転を減らす馬力制御傾転アクチュエータ１２ａと、油圧ポンプ２の吐出圧が複数のアクチュエータ３ａ，３ｂ，３ｃの最高負荷圧より目標差圧だけ高くなるようロードセンシング制御するＬＳ制御弁１２ｂ及びＬＳ制御傾転アクチュエータ１２ｃとを備えている。
【００３０】
ＬＳ制御弁１２ｂは、アクチュエータ１２ｃを増圧し油圧ポンプ２の傾転を減らす側に位置する受圧部１２ｄと、アクチュエータ１２ｃを減圧し油圧ポンプ２の傾転を増やす側に位置する受圧部１２ｅとを有し、受圧部１２ｄには差圧減圧弁１１の出力圧（油圧ポンプ２の吐出圧とアクチュエータ３ａ，３ｂ，３ｃの最高負荷圧との差圧）が導かれ、受圧部１２ｅにはエンジン回転数検出回路１３の差圧減圧弁５１の出力圧がロードセンシング制御の目標差圧（目標ＬＳ差圧）として導かれている。
【００３１】
また、ポンプ傾転制御機構１２はポンプ傾転可変機構９０を備え、アクチュエータ１２ａ，１２ｃはこのポンプ傾転可変機構９０を作動することで油圧ポンプ２の傾転を制御する。ポンプ傾転可変機構９０は、油圧ポンプ２の最少吐出流量を旋回最大要求流量以上に設定するよう油圧ポンプ２の傾転を制限するリミッタ９１を備えている（後述）。
【００３２】
エンジン回転数検出回路１３は、流量検出弁５０と上記の差圧減圧弁５１とを有し、流量検出弁５０は可変の絞り部５０ａを有しかつその絞り部５０ａがパイロットポンプ３０の吐出ライン３１に配置されている。吐出ライン３１は流量検出弁５０の上流側のライン３１ａと下流側のライン３１ｂを有し、下流側のライン３１ｂには、パイロット油圧源としての元圧を規定するリリーフ弁３２が接続され、ライン３１ｂは、例えば流量制御弁６ａ，６ｂ，６ｃを切換操作するためのパイロット圧を生成するリモコン弁（図示せず）へと接続されている。
【００３３】
流量検出弁５０は吐出ライン３１を流れる圧油の流量を絞り部５０ａの前後差圧の変化として検出し、その前後差圧を目標ＬＳ差圧として用いるためのものである。ここで、吐出ライン３１を流れる圧油の流量はパイロットポンプ３０の吐出流量であり、この吐出流量はエンジン１の回転数によって変化するため、吐出ライン３１を流れる圧油の流量を検出することはエンジン１の回転数を検出することである。例えば、エンジン１の回転数が低下すれば当該流量が減少し、絞り部５０ａの前後差圧は低下する。
【００３４】
また、絞り部５０ａは開口面積が連続的に変化する可変絞り部として構成されており、流量検出弁５０は更に開方向作動の受圧部５０ｂと絞り方向作動の受圧部５０ｃ及びバネ５０ｄを有し、受圧部５０ｂに可変絞り部５０ａの上流側圧力（ライン３１ａの圧力）が導かれ、受圧部５０ｃに可変絞り部５０ａの下流側圧力（ライン３１ｂの圧力）が導かれ、可変絞り部５１ａ自身の前後差圧に依存してその開口面積を変化させる構成となっている。このように流量検出弁５０を構成し、可変絞り部５０ａの前後差圧を目標ＬＳ差圧として用いることにより、エンジン回転数に応じたサチュレーション現象の改善が図れ、エンジン回転数を低く設定した場合に良好な微操作性が得られる。なお、この点は特開平１０−１９６６０４号公報に詳しい。
【００３５】
差圧減圧弁５１は、エンジン回転数に依存する圧力として可変絞り部５０ａの前後差圧を絶対圧として出力するエンジン回転数検出弁であり、出力圧を増やす側に位置する受圧部５１ａと出力圧を減らす側に位置する受圧部５１ｂ，５１ｃを有し、受圧部５１ａに可変絞り部５０ａの上流側圧力が導かれ、受圧部５１ｂ，５１ｃにそれぞれ可変絞り部５０ａの下流側圧力と自己の出力圧が導かれ、これらの圧力のバランスでライン３１ｂの圧力を基に可変絞り部５０ａの前後差圧を絶対圧として出力する。
【００３６】
差圧減圧弁５１の出力ポートは信号ライン５３を介してＬＳ制御弁１２ｂの受圧部１２ｅに接続され、差圧減圧弁５１の出力圧が目標ＬＳ差圧として受圧部１２ｅに導かれる。その結果、エンジン回転数に応じたアクチュエータスピードの設定が可能となる。また、差圧減圧弁５１の出力ポートは信号ライン５３，５４を介して圧力補償弁７ａの受圧部７０ｃに接続され、差圧減圧弁５１の出力圧が上記のように目標補償差圧として受圧部７０ｃに導かれる。
【００３７】
図２はポンプ傾転制御機構１２による油圧ポンプ２のＰＱ（圧力・流量）特性図である。横軸は油圧ポンプ２の吐出圧力、縦軸は油圧ポンプ２の吐出流量であり、特性線Ｈは馬力制御傾転アクチュエータ１２ａによるポンプ馬力制御線図である。また、Ｑminは油圧ポンプ２の最少吐出流量であり、Ｑsmaxは旋回最大要求流量である。旋回最大要求流量Ｑsmaxは、旋回流量制御弁６ａのメータイン絞り部６１又は６２の最大開口面積をＡmaxとすると、Ｑsmax＝Ｃ・Ａmax√Ｐgr（Ｃは係数）で表される。つまり、一般的に要求流量は開口面積とその前後差圧の関数であり、流量制御弁６ａの前後差圧は圧力補償弁７ａによりその目標補償差圧に等しくなるよう制御され、圧力補償弁７ａの受圧部７０ｃには、その目標補償差圧として差圧減圧弁５１の出力圧が導かれているので、差圧減圧弁５１の出力圧をＰ gr とすると、Ｑsmax＝Ｃ・Ａmax√Ｐgrとなる。
【００３８】
上述したように、油圧ポンプ２は、リミッタ９１により最少吐出流量Ｑminが旋回最大要求流量Ｑsmax以上になるように制限されている。このため、ＬＳ制御弁１２ｂ及びＬＳ制御傾転アクチュエータ１２ｃによるポンプＬＳ制御領域は斜線部分となる。
【００３９】
次に、本実施の形態の動作を比較例と対比して説明する。以下の説明では、油圧ポンプ２の吐出圧をＰｓ、アクチュエータ３ａ，３ｂ，３ｃの最高負荷圧をＰLmax、油圧ポンプ２の吐出圧と最高負荷圧との差圧（ＬＳ差圧）をΔＰLS、差圧減圧弁１１の出力圧をＰLS、差圧減圧弁５１の出力圧をＰgr、圧力補償弁７ａの目標補償差圧をＰc1、圧力補償弁７ｂの目標補償差圧をＰc2、圧力補償弁７ｃの目標補償差圧をＰc3とする。
【００４０】
図３は、比較例１として、図１に示した旋回セクションの圧力補償弁７ａの受圧部７０ｃに差圧減圧弁５１の出力圧でなく、他のセクションの圧力補償弁７ｂ，７ｃと同様に差圧減圧弁１１の出力圧を導くと共に、ポンプ傾転制御機構１２Ｘのポンプ傾転可変機構９０にリミッタ９１を設けず、油圧ポンプ２の最少吐出流量を通常の値に設定したものである。この構成は、特開平１０−１９６６０４号公報に記載の従来技術に差圧減圧弁１１を付加し、圧力補償弁７ａ，７ｂ，７ｃに油圧ポンプ２の吐出圧Ｐｓと複数のアクチュエータ３ａ，３ｂ，３ｃの最高負荷圧ＰLmaxとの差圧ΔＰLSを絶対圧（差圧減圧弁１１の出力圧ＰLS）として導くようにした構成に相当する。
【００４１】
図３の比較例１においては、１つのアクチュエータ、例えば旋回モータ３ａを駆動する単独動作から複数のアクチュエータ、例えば旋回モータ３ａとブームシリンダ３ｂを同時に駆動する複合動作に移行する場合に、操作性の低下が発生する。
【００４２】
複数のアクチュエータ３ａ，３ｂを同時に駆動する複合動作において、駆動している全てのアクチュエータに対する要求流量の総和が油圧ポンプ２の吐出可能な最大流量を越えた場合、油圧ポンプ２の吐出流量が不足するサチュレーション状態が発生する。
【００４３】
この場合、流量制御弁４ａ，４ｂにおいては、それぞれのアクチュエータ３ａ，３ｂヘの流量の再分配が行われる。つまり、アクチュエータ３ａ，３ｂに係わる圧力補償弁７ａ，７ｂの目標補償差圧Ｐc1，Ｐc2は油圧ポンプ２の吐出圧Ｐｓと最高負荷圧ＰLmaxの差圧（ＬＳ差圧）ΔＰLS＝Ｐｓ‐ＰLmaxで与えられる。そのＬＳ差圧は差圧減圧弁１１から絶対圧ＰＬＳとして供給される。サチュレーション状態では、ＬＳ差圧ΔＰLSは供給不足の比率に応じて低下する。このため、その差圧により設定される圧力補償弁７ａ，７ｂの目標補償差圧Ｐc1，Ｐc2も、本来は目標ＬＳ差圧と同じになるべきものがそれよりも供給不足の比率分減少する。その結果、各アクチュエータには、油圧ポンプ２の吐出流量をそれぞれの要求流量比に分けた流量が供給される。これにより各流量制御弁６ａ，６ｂに対する操作レバーの入力比に応じて全体のアクチュエータスピードが低下することになる。この機能により、複合動作時に、特定のアクチュエータが停止してしまうことなく、操作性を維持することが可能になる。
【００４４】
しかし、この複合動作において、全てのアクチュエータの速度が均等に低下するため、操作性が低下する場合がある。
【００４５】
例えば、旋回操作ではオペレータが速度変化を体感できるため、速度変化に敏感であり、他のアクチュエータに比べ速度変化を抑えた構成とする必要がある。また、吊り荷作業等でも、旋回中に速度変化が生じると、荷がオペレータの意図に反して揺れてしまう。
【００４６】
また、特開２０００−２２７１０３号公報のように、旋回優先バネを圧力補償弁７ａに設定する方法も考えられる。しかし、この場合はエンジン回転数によって目標補償差圧が変化するため、全てのエンジン回転数において、同じ割合で優先性を保持することが不可能である。
【００４７】
そこで、旋回セクションの圧力補償弁７ａの受圧部７０ｃにＬＳ制御の目標値（目標ＬＳ差圧）である差圧減圧弁５１の出力圧Ｐgrを導き、その目標補償差圧Ｐc1を出力圧Ｐgrにより設定することで、サチュレーションの情報が旋回圧力補償弁７ａに伝わらないようにする方法が考えられる。図４は、その場合の構成を比較例２として示すものである。
【００４８】
しかし、上記構成では、旋回の単独動作時に、旋回セクションの圧力補償弁７ａとポンプ傾転制御機構１２が、共に旋回セクションのメインスプールである流量制御弁６ａのメータイン絞り６１又は６２の前後差圧を差圧減圧弁５１の出力圧Ｐgrに制御するよう動作し、両者が干渉してしまう。
【００４９】
以下、そのメカニズムについて説明する。
【００５０】
図５はポンプ傾転制御機構１２Ｘによる、図２と同様な油圧ポンプ２のＰＱ（圧力・流量）特性図である。Ｑmin0はリミッタ９１のない場合の通常の油圧ポンプ２の最少吐出流量である。この油圧ポンプ２の最少吐出流量Ｑmin0は旋回最大要求流量Ｑsmaxより少なく、ＬＳ制御弁１２ｂ及びＬＳ制御傾転アクチュエータ１２ｃによるポンプＬＳ制御領域に旋回最大要求流量Ｑsmaxが包含されている。
【００５１】
旋回の単独動作において、旋回の起動直後から加速し、目標速度に達し加速がなくなると、旋回負荷圧の低下により、油圧ポンプ２の吐出圧Ｐｓと最高負荷圧ＰLmaxの差圧ΔＰLSが、上記ＬＳ制御の目標値である差圧減圧弁５１の出力圧Ｐgrより大きくなる。その結果、ポンプ傾転制御機構１２は、ポンプ供給油路５のポンプ吐出圧Ｐｓを下げる方向に制御し、油圧ポンプ２の傾転は小さくなり、吐出流量は減少する。これに対して、旋回セクションの圧力補償弁７ａは流量制御弁６ａのメータイン絞り６１又は６２の前後差圧を下げるために閉じる方向に動き、ポンプ供給油路５の圧油が閉じこめられポンプ吐出圧Ｐｓが下がりにくくなる。つまり、油圧ポンプ２のＬＳ制御は吐出流量を減らしポンプ吐出圧Ｐｓ下げるように制御するのに対して、圧力補償弁７ａが閉じ方向に動き、油圧ポンプ２の吐出圧Ｐｓを下がりにくくするように制御する。ここで、油圧ポンプ２の最少吐出流量Ｑmax0は旋回最大要求流量Ｑsmaxより少ない値である。その結果、油圧ポンプ２の吐出流量はＬＳ制御により極端に低下する。このように油圧ポンプ２の動作と旋回圧力補償弁７ａの動作は干渉し、油圧ポンプ２の吐出圧の制御性は低下する。
【００５２】
また、その後、旋回モータ３ａの回転により油圧ポンプ２の吐出流量が消費されると、油圧ポンプ２の吐出流量が極端に低下しているために供給不足になる。その結果、油圧ポンプ２の吐出圧Ｐｓと最高負荷圧ＰLmaxの差圧ΔＰLSが上記ＬＳ制御の目標値であるＰgrより低くなる。このとき、旋回セクションの圧力補償弁７ａが全開となり機能しなくなる。これと同時にポンプ傾転制御機構１２は、ポンプ吐出圧Ｐｓと最高負荷圧ＰLmaxの差圧ΔＰLSがＬＳ制御の目標値であるＰgrより大きくなる方向、つまり油圧ポンプ２の吐出圧を上げる方向に制御し、油圧ポンプ２の傾転は大きくなり、吐出流量は増加する。その際、旋回セクションの圧力補償弁７ａが全開しているため、油圧ポンプ２の吐出圧Ｐｓは上がり難く、制御の応答性が悪化し、油圧ポンプ２の吐出流量はＬＳ制御により極端に上昇する。
【００５３】
以上の現象の繰り返しにより、不連続な加速が生じ、操作性を著しく低下してしまう。この現象は、慣性負荷を伴う旋回に顕著に見られる。
【００５４】
このような問題を解決するため、本実施の形態では、上記のように旋回セクションの圧力補償弁７ａの受圧部７０ｃに差圧減圧弁５１の出力圧を導き、かつ油圧ポンプ２の最少吐出流量を旋回最大要求流量よりも高く設定したものである。
【００５５】
図６は、本実施の形態における圧力補償弁７ａ，７ｂ，７ｃの目標補償差圧Ｐc1，Ｐc2，Ｐc3の変化を示す図である。図中Ａは旋回単独動作時（非サチュレーション状態）、Ｂは旋回複合動作時（サチュレーション状態）であり、共にエンジン１が定格の高回転数にあるときのもの、Ｃは旋回複合動作時（サチュレーション状態）でエンジン１が低速回転数にあるときのものである。
【００５６】
旋回セクションの圧力補償弁７ａの目標補償差圧Ｐc1は、ＬＳ制御の目標値である差圧減圧弁５１の出力圧Ｐgrであり、それ以外の圧力補償弁７ｂ，７ｃの目標補償差圧Ｐc2，Ｐc3は、差圧減圧弁１１の出力圧ＰLS（ＬＳ差圧ΔＰLSの絶対圧）となるよう制御される。
【００５７】
まず、エンジン１が定格の高回転数に保たれる場合は、差圧減圧弁５１の出力圧Ｐgrは変化せず、旋回単独動作時も複合動作時（サチュレーション状態）も旋回セクションの圧力補償弁７ａの目標補償差圧Ｐc1は一定である（図６のＡ及びＢ）。このため、旋回単独動作時で油圧ポンプの吐出流量が足りているときは（図６のＡ）、差圧減圧弁１１の出力圧ＰLSはＰgrに等しくなるよう制御されるため（ＰLS＝Ｐgr）、Ｐgrを目標補償差圧としている旋回と、ＰLSを目標としている他のアクチュエータとでは、コントロールバルブ４の圧力補償弁７ａ，７ｂ，７ｃによる制御動作には基本的な差はない。
【００５８】
旋回を含んだ複合操作（サチュレーション状態）では、コントロールバルブ４において、それぞれのアクチュエータヘの流量の再分配が行われる。
【００５９】
まず、旋回セクション以外の圧力補償弁７ｂ，７ｃの目標補償差圧Ｐc2，Ｐc3は、サチュレーション状態になるとポンプ吐出流量の供給不足の比率に応じて低下する。つまり、圧力補償弁７ｂ，７ｃの目標補償差圧Ｐc2，Ｐc3は油圧ポンプ２の吐出圧Ｐｓと最高負荷圧ＰLmaxの差圧ΔＰLS（＝Ｐｓ−ＰLmax）で、差圧減圧弁１１により絶対圧ＰLSとして与えられる。サチュレーション状態では、ＬＳ差圧ΔＰLSは供給不足の比率に応じて低下する。このため、図６のＢに示すように、その差圧により設定される圧力補償弁７ｂ，７ｃの目標補償差圧Ｐc2，Ｐc3もそれに応じて減少する。その結果、旋回を除く各アクチュエータ３ｂ，３ｃには、油圧ポンプ２の吐出流量をそれぞれの要求流量比に分けた流量が供給される。これにより各流量制御弁６ｂ，６ｃに対する操作レバーの入力比に応じて全体のアクチュエータスピードが低下することになる。
【００６０】
しかし、旋回に関しては、圧力補償弁７ａの目標補償差圧Ｐc1をＬＳ差圧でなく差圧減圧弁５１の出力圧Ｐgrにより設定しており、このＰgrはサチュレーション状態であってもエンジン回転数が定格の高回転数にある限り変わらず（図６のＢ）、サチュレーションの情報が圧力補償弁７ａに伝わらない。その結果、サチュレーション状態でのＰc1の低下が抑えられ、他のアクチュエータ３ｂ，３ｃに比べ旋回モータ３ａに優先的に圧油を供給し得るようになり、旋回を含んだ複合動作による旋回モータ３ａのスピードダウンを抑えることが可能になる。
【００６１】
また、エンジン回転数を低速に下げた場合は、差圧減圧弁５１の出力圧Ｐgrがそれに応じて減少するため、圧力補償弁７ａの目標補償差圧Ｐc1（＝Ｐgr）も同様に減少する。旋回セクション以外の圧力補償弁７ｂ，７ｃの目標補償差圧Ｐc2，Ｐc3も、エンジン１が低速回転数になると油圧ポンプ２の吐出圧Ｐｓと最高負荷圧ＰLmaxの差圧ΔＰLS（＝Ｐｓ−ＰLmax）が差圧減圧弁５１の出力圧Ｐgrに応じて低下するよう制御されるため、目標補償差圧Ｐc2，Ｐc3も減少する。つまり、図６のＣに示すように、エンジン回転数を低速にしたときの旋回複合操作（サチュレーション状態）では、圧力補償弁７ａの目標補償差圧Ｐc1と圧力補償弁７ｂ，７ｃの目標補償差圧Ｐc2，Ｐc3が共に減少する。このため旋回が速くなりすぎるというようなことが起こらず、エンジン回転数の設定によらず、適切な旋回の優先性を維持することができる。
【００６２】
次に、旋回単独操作の起動時から定常状態に移行するときの動作について説明する。
【００６３】
旋回起動時に、油圧ポンプ２の吐出圧Ｐｓと最高負荷圧ＰLmaxの差圧ΔＰLSがＬＳ制御の目標値であるＰgrより小さくなる。したがって、油圧ポンプ２はポンプ供給油路５のポンプ吐出圧Ｐｓを上げる方向に制御し、ポンプ傾転は大きくなり、吐出流量は増加する。同時に旋回セクションの圧力補償弁７ａは、メータイン絞り部６１又は６２の前後差圧（＝ΔＰLS）が目標補償差圧Ｐc1（＝Ｐgr）より小さいため、全開となり機能しなくなる。
【００６４】
旋回速度が目標値になり加速がなくなると、油圧ポンプ２の吐出圧Ｐｓと最高負荷圧Ｐ１maxの差圧ΔＰLSがＬＳ制御の目標値であるＰgrより大きくなる。油圧ポンプ２はポンプ供給油路５のポンプ吐出圧Ｐｓを下げる方向に制御し、油圧ポンプ２の傾転は小さくなり、吐出流量は減少する。同時に旋回セクションの圧力補償弁７ａは閉じる方向に動作し、メータイン絞り部６１又は６２の前後差圧を目標補償差圧Ｐc1（＝Ｐgr）となるよう制御する。
【００６５】
ここで、油圧ポンプ２は、リミッタ９１により最少吐出流量Ｑminが旋回最大要求流量Ｑsmax以上になるように制限されている。このため、油圧ポンプ２の吐出流量が最少吐出流量Ｑminに達すると、それ以降はＱminに維持され、油圧ポンプ２のＬＳ制御は動作しなくなる。つまり、油圧ポンプ２は容量Ｑminの固定ポンプであるかの如く動作する。その結果、油圧ポンプ２のＬＳ制御と旋回圧力補償弁７ａの干渉はなくなる。
【００６６】
図７及び図８は、以上の説明のうち、図４に示した比較例２と図１に示した本発明における旋回操作時にポンプと旋回圧力補償弁の動きを、図３に示した比較例１と対比し、ＬＳ差圧ΔＰLS（＝Ｐｓ−ＰLmax）がＬＳ制御の目標値であるＰgrより小さくなる場合と、大きくなる場合とについて表形式でまとめたものである。
【００６７】
以上のように本実施の形態によれば、ＬＳシステムを備えた油圧駆動装置において、ポンプ吐出流量のサチュレーション状態が生じても特定のアクチュエータ３ａに優先的に圧油を供給してその速度変化を抑えることができ、かつエンジン回転数の設定に係わらず同様に優先性を維持することができ、優れた操作性を実現することができる。
【００６８】
また、油圧ポンプ２のロードセンシング制御と圧力補償弁７ａ〜７ｃの動作の干渉を回避することができ、システムの安定性を維持することが可能になる。
【００６９】
なお、本実施の形態では、油圧ポンプ２の吐出圧と最高負荷圧との差圧を差圧減圧弁１１の出力圧により絶対圧として導き、エンジン検出回路１３の可変絞り部５０ａの前後差圧を差圧減圧弁５１の出力圧により絶対圧として導くようにしたが、油圧ポンプ２の吐出圧と最高負荷圧、可変絞り部５０ａの上流側圧力と下流側圧力をそれぞれ別々に導くようにしてもよい。
【００７０】
【発明の効果】
本発明によれば、ＬＳシステムを備えた油圧駆動装置において、ポンプ吐出流量のサチュレーション状態が生じても特定のアクチュエータに優先的に圧油を供給してその速度変化を抑えることができ、かつエンジン回転数の設定に係わらず同様に優先性を維持することができ、優れた操作性を実現することができる。
【００７１】
また、油圧ポンプのロードセンシング制御と圧力補償弁の動作の干渉を回避することができ、システムの安定性を維持することが可能になる。
【図面の簡単な説明】
【図１】本発明の一実施の形態に係わる油圧駆動装置の全体構成を示す図である。
【図２】ポンプ傾転制御機構による油圧ポンプのＰＱ特性を示す図である。
【図３】従来技術に基づく構成を比較例１として示す図１と同様な図である。
【図４】比較例１の対策例を比較例２として示す図１と同様な図である。
【図５】比較例１及び２におけるポンプ傾転制御機構による油圧ポンプのＰＱ特性を示す図である。
【図６】図１に示した油圧駆動装置における目標補償差圧の変化を示す図である。
【図７】比較例２の旋回動作時のポンプと旋回圧力補償弁の動きを表形式で示す図である。
【図８】本実施の形態の旋回動作時のポンプと旋回圧力補償弁の動きを表形式で示す図である。
【符号の説明】
１エンジン
２メインの油圧ポンプ
３ａアクチュエータ（旋回モータ９）
３ｂアクチュエータ（ブームシリンダ）
３ｃアクチュエータ（アームシリンダ）
４コントロールバルブ
４ａ，４ｂ，４ｃ流量制御弁
５供給油路
６ａ，６ｂ，６ｃ流量制御弁
７ａ，７ｂ，７ｃ圧力補償弁
８ａ，８ｂ，８ｃ、８ｄ負荷ライン
９ａ，９ｂシャトル弁
１０信号ライン
１１差圧減圧弁
１１ａ，１１ｂ，１１ｃ受圧部
１２ポンプ傾転制御機構
１２ａ馬力制御傾転アクチュエータ
１２ｂＬＳ制御弁
１２ｃＬＳ制御傾転アクチュエータ
１３エンジン回転数検出回路
３１吐出ライン
３２リリーフ弁
５０流量検出弁
５０ａ絞り部
５１差圧減圧弁
５１ａ，５１ｂ，５１ｃ受圧部
５３，５４信号ライン
６０ａ，６０ｂ，６０ｃ負荷ポート
６１，６２メータイン可変絞り部
７０ａ，７０ｂ，７０ｃ受圧部
７１ａ，７１ｂ，７１ｃ受圧部
７２ａ，７２ｂ，７２ｃ受圧部
９０ポンプ傾転可変機構
９１リミッタ[0001]
BACKGROUND OF THE INVENTION
The present invention relates to an LS control type hydraulic drive device used for a hydraulic excavator, and in particular, performs load sensing control so that a discharge pressure of a hydraulic pump is higher than a maximum load pressure of a plurality of actuators by a target differential pressure, and a plurality of flow rates. Each target compensation differential pressure of the pressure compensation valve that controls the differential pressure across the control valve is set by the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of multiple actuators, and the target differential pressure for load sensing control The present invention relates to a hydraulic drive device for a hydraulic excavator that is set as a variable value that depends on the engine speed.
[0002]
[Prior art]
A hydraulic drive device that performs load sensing control so that the discharge pressure of the hydraulic pump is higher than the maximum load pressure of a plurality of actuators by a target differential pressure is called a load sensing system (hereinafter referred to as LS system as appropriate). The front and rear differential pressures are controlled by pressure compensation valves, respectively, so that pressure oil can be supplied at a ratio corresponding to the opening area of the flow control valve regardless of the load pressure at the time of combined operation in which a plurality of actuators are driven simultaneously.
[0003]
In such an LS system, a differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of a plurality of actuators (hereinafter referred to as LS differential pressure) is led to the pressure compensation valve, and each target compensation differential pressure of the pressure compensation valve is calculated. It is generally performed by setting the LS differential pressure. This makes it possible to maintain a saturation state in which the discharge flow rate of the hydraulic pump does not satisfy the flow rate required by the multiple flow control valves during the combined operation of simultaneously driving the multiple actuators. Even if the LS differential pressure decreases, the LS differential pressure decreases according to the degree of saturation, and the target compensation differential pressure of the pressure compensation valve also decreases accordingly. Therefore, the discharge flow rate of the hydraulic pump is the ratio of the flow rate required by each actuator. Can be redistributed.
[0004]
In such an LS system, Japanese Patent Application Laid-Open No. 10-89304 is provided with a differential pressure reducing valve that outputs a differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of a plurality of actuators as an absolute pressure. There is a description in which the output pressure of the pressure reducing valve is guided to a plurality of pressure compensating valves, and each target compensating differential pressure is set.
[0005]
Japanese Patent Application Laid-Open No. 10-196604 is provided with a differential pressure reducing valve that outputs, as an absolute pressure, a pressure that depends on the number of revolutions of an engine that drives a hydraulic pump, and the output pressure of the differential pressure reducing valve is set as an LS control regulator. Thus, there is described a hydraulic drive apparatus in which a target differential pressure of load sensing control (hereinafter referred to as a target LS differential pressure) is set as a variable value depending on the engine speed.
[0006]
Further, Japanese Patent Laid-Open No. 2000-227103 discloses that the pressure compensation valve of the swing section has a load-dependent characteristic that reduces the target compensation differential pressure when the swing negative pressure increases, so that the swing motor can be smoothly accelerated and steady. The lower limit setting spring (swing priority spring) that prevents the target compensation differential pressure from falling below the specified value is provided, and even if the pump discharge flow rate is in a saturation state at the start of the swing combined operation, A hydraulic drive device is described in which the turning speed is not extremely slow compared to other actuators.
[0007]
[Problems to be solved by the invention]
However, the above-described conventional technique has a problem that operability is deteriorated when shifting from a single operation for driving one actuator to a composite operation for simultaneously driving a plurality of actuators.
[0008]
That is, in a combined operation in which a plurality of actuators are driven simultaneously, a saturation state occurs when the sum of the required flow rates of all the operating actuators exceeds the maximum flow rate that can be discharged by the hydraulic pump. In this state, the LS differential pressure decreases according to the degree of saturation, and the target compensation differential pressure of the pressure compensation valve decreases accordingly. Therefore, each actuator requires the discharge flow rate of the hydraulic pump in each flow control valve. Can be redistributed to the ratio of the flow rate to be. As a result, the overall actuator speed decreases according to the input ratio of the operation lever to each flow control valve, and it is possible to maintain operability without stopping specific actuators during combined operation. Become.
[0009]
However, in this combined operation, the speed of all the actuators is reduced uniformly, so that the operability may be reduced.
[0010]
For example, in a turning operation of a hydraulic excavator, an operator can feel a change in speed, so that the operator is sensitive to the change in speed and needs to have a configuration in which the change in speed is suppressed compared to other actuators. In addition, even in a suspended load operation or the like, if a speed change occurs during turning, the load will sway against the operator's intention.
[0011]
As disclosed in Japanese Patent Application Laid-Open No. 2000-227103, a method of setting the swing priority spring as a pressure compensation valve is also conceivable. However, in this case, since the target compensation differential pressure varies depending on the engine speed, it is impossible to maintain the priority at the same rate at all engine speeds.
[0012]
An object of the present invention is to provide a hydraulic drive device for a hydraulic excavator equipped with an LS system, so that even if a saturation state of the pump discharge flow rate occurs, pressure oil is preferentially supplied to the swing motor, and the speed change thereof can be suppressed. In addition, it is possible to provide a device that can maintain priority in the same manner regardless of the setting of the engine speed and can realize excellent operability.
[0013]
[Means for Solving the Problems]
  (1) In order to achieve the above object, the present invention includes a plurality of engines, a variable displacement hydraulic pump driven by the engine, and a swing motor driven by pressure oil discharged from the hydraulic pump. A plurality of flow rate control valves that control flow rates of pressure oil supplied from the hydraulic pump to the plurality of actuators, and a plurality of pressure compensation valves that respectively control the differential pressure across the plurality of flow rate control valves, A pump control means for performing load sensing control so that a discharge pressure of the hydraulic pump is higher than a maximum load pressure of the plurality of actuators by a target differential pressure;So that the engine speed decreases as the engine speed decreasesAn engine speed detection circuit that outputs a pressure depending on the engine speed, and the output pressure of the engine speed detection circuit is led to the pump control means,By setting the target differential pressure of the load sensing control by the output pressure of the engine speed detection circuit, the target differential pressure of the load sensing control decreases as the engine speed decreases.In the hydraulic drive device of a hydraulic excavator in which the target differential pressure of the load sensing control is set as a variable value that depends on the number of revolutions of the engine, the target compensated differential pressure of each pressure compensation valve related to an actuator other than the swing motor is The engine rotational speed detection circuit is set by the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators, and the output pressure of the engine rotational speed detection circuit is led to a pressure compensation valve related to the swing motor. The target compensation differential pressure is set by the output pressure of the pump, and the pump control means,in frontIt is assumed that the hydraulic pump has a variable pump tilt mechanism including a limiter means for limiting the tilt of the hydraulic pump so that the minimum discharge flow rate of the hydraulic pump does not become smaller than the maximum required flow rate of the flow control valve related to the swing motor.
[0014]
Thus, each target compensation differential pressure of the pressure compensation valve related to the actuator other than the swing motor is set by LS differential pressure, and the target compensation differential pressure of the pressure compensation valve related to the swing motor is set as the output pressure of the engine speed detection circuit ( In the combined operation including the swing motor (saturation state), in the actuator other than the swing motor, the target compensation differential pressure of the pressure compensation valve is set according to the decrease in the LS differential pressure. However, saturation information is not transmitted to the pressure compensation valve on the swing motor side, and the target compensation differential pressure remains the same as when the saturation state is not reached. Is supplied and it is possible to suppress the speed reduction.
[0015]
Also, when the engine speed is lowered, the output pressure of the engine speed detection circuit (target differential pressure for load sensing control) decreases accordingly, and the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure is Not only the target compensation differential pressure of the pressure compensation valve related to the actuator other than the swing motor is reduced, but also the target compensation differential pressure of the pressure compensation valve related to the swing motor is similarly reduced. As a result, the turning motor does not become too fast, and appropriate turning priority can be maintained regardless of the setting of the engine speed.
[0016]
  Furthermore, a pump tilt variable mechanism is provided in the pump control means, and this pump tilt variable mechanism,oilBy providing a limiter means to limit the tilting of the hydraulic pump so that the minimum discharge flow rate of the pressure pump does not become smaller than the maximum required flow rate of the flow control valve related to the swing motor, the actuator speed can be set as the target when a specific actuator is operated independently. When load sensing control is performed to reach the speed and the hydraulic pump discharge flow rate decreases, when the hydraulic pump discharge flow rate reaches the minimum discharge flow rate, the minimum discharge flow rate is maintained thereafter, and the LS control of the hydraulic pump operates. Disappear. In other words, the hydraulic pump operates as if it were a fixed pump with the minimum discharge flow rate as the capacity. As a result, the interference between the LS control of the hydraulic pump and the swing pressure compensation valve can be avoided, and the stability of the system can be maintained.
[0018]
  (2In the above (1), preferably, the engine rotation speed detection circuit includes a first differential pressure reducing valve that outputs a pressure depending on the rotation speed of the engine as an absolute pressure,Slewing motorThe output pressure of the first differential pressure reducing valve is introduced to the pressure compensating valve related to the other actuators, and the respective target compensating differential pressures are set.
[0019]
  (3) Further, in the above (1), preferably further comprising a second differential pressure reducing valve that outputs a differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators as an absolute pressure,Slewing motorThe output pressure of the second differential pressure reducing valve is led to the pressure compensation valve related to the output of the second differential pressure reducing valve.With pressureSet the target compensation differential pressure of the pressure compensation valve.
[0020]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
[0021]
FIG. 1 is a view showing a hydraulic drive apparatus according to an embodiment of the present invention.
[0022]
In FIG. 1, a hydraulic drive apparatus according to the present embodiment includes an engine 1, a variable displacement hydraulic pump 2 and a fixed displacement pilot pump 30 as main pumps driven by the engine 1, and a main hydraulic pressure. Pressure oil connected to the plurality of actuators 3a, 3b, 3c driven by the pressure oil discharged from the pump 2 and the supply oil passage 5 of the hydraulic pump 2 and supplied from the hydraulic pump 2 to the actuators 3a, 3b, 3c The differential pressure (LS differential pressure) between the discharge pressure of the plurality of flow control valves (main spools) 4a, 4b, 4c and the hydraulic pump 2 and the maximum load pressure of the plurality of actuators 3a, 3b, 3c that respectively control the flow rate and direction of ), And a pump tilt control mechanism 1 that controls the tilt (capacity) of the hydraulic pump 2. If, and an engine rotation speed detection circuit 13 including the differential pressure reducing valve 51 to output a pressure that depends on the engine speed as an absolute pressure.
[0023]
The hydraulic drive apparatus according to the present embodiment is mounted on, for example, a hydraulic excavator, and the actuators 3a, 3b, 3c are, for example, a swing motor, a boom cylinder, and an arm cylinder of the hydraulic excavator, respectively. The hydraulic excavator has an upper swing body that is pivotably mounted on the lower traveling body, and a front working mechanism that includes a boom, an arm, and a bucket that is mounted on the upper swing body so as to be pivotable in the vertical direction. The upper swing body is driven to rotate, and the boom cylinder 3b and the arm cylinder 3c rotate the boom and arm in the vertical direction, respectively.
[0024]
The plurality of flow control valves 4a, 4b, and 4c are respectively a plurality of closed center type flow control valves 6a, 6b, and 6c and before and after meter-in throttle portions 61 and 62 of the plurality of flow control valves 6a, 6b, and 6c. It comprises a plurality of pressure compensating valves 7a, 7b, 7c that control the differential pressure to the same value.
[0025]
The flow rate control valves 6a, 6b, and 6c are each switched by operating an operation lever (not shown), and the opening area of the meter-in throttle 61 or 62 is determined according to the operation amount of the operation lever.
[0026]
The plurality of pressure compensation valves 7a, 7b, and 7c are the front type (before orifice type) installed upstream of the meter-in throttle portions 61 and 62 of the flow control valves 6a, 6b, and 6c, respectively. A pair of opposed pressure receiving portions 70a and 70b and an opening direction pressure receiving portion 70c are provided, and pressures on the upstream side and downstream side of the flow rate control valve 6a are guided to the pressure receiving portions 70a and 70b, respectively. The output pressure of the differential pressure reducing valve 51 is guided, and the differential pressure across the flow rate control valve 6a is controlled using the output pressure as the target compensation differential pressure. The pressure compensating valve 7b has a pair of opposed pressure receiving portions 71a and 71b and an opening direction pressure receiving portion 71c, and the pressures on the upstream side and the downstream side of the flow control valve 6b are respectively introduced to the pressure receiving portions 71a and 71b. Then, the output pressure of the differential pressure reducing valve 11 is guided to the pressure receiving portion 71c, and the front-rear differential pressure of the flow control valve 6b is controlled using the output pressure as the target compensation differential pressure. The pressure compensation valve 7c is the same as the pressure compensation valve 7b, and has a pair of opposed pressure receiving portions 72a and 72b and an opening direction pressure receiving portion 72c. The pressure receiving portions 72a and 72b are upstream of the flow rate control valve 6c. Then, the pressure on the downstream side is guided, the output pressure of the differential pressure reducing valve 11 is guided to the pressure receiving portion 72c, and the differential pressure across the flow control valve 6c is controlled using the output pressure as the target compensation differential pressure.
Each of the plurality of flow control valves 6a, 6b, 6c is provided with load ports 60a, 60b,... For taking out the load pressures when the actuators 3a, 3b, 3c are driven, and these load ports 60a, 60b,. The highest pressure among the extracted load pressures is detected on the signal line 10 via the load lines 8a, 8b, 8c, 8d and the shuttle valves 9a, 9b.
[0027]
The differential pressure reducing valve 11 has a pressure receiving portion 11a positioned on the side that increases the output pressure and pressure receiving portions 11b and 11c positioned on the side that decreases the output pressure, and the discharge pressure of the hydraulic pump 2 is guided to the pressure receiving portion 11a. The maximum load pressure detected in the signal line 10 and its own output pressure are guided to the pressure receiving portions 11b and 11c, respectively, and the differential pressure (LS differential pressure) between the discharge pressure of the hydraulic pump 2 and the maximum load pressure is balanced by these pressures. ) Is output as absolute pressure.
[0028]
The output port of the differential pressure reducing valve 11 is connected to the pressure receiving portion 12d of the LS control valve 12b provided in the pump tilt control mechanism 12 via the signal lines 21 and 22, and the output pressure of the differential pressure reducing valve 11 is the pressure receiving portion. 12d. The output port of the differential pressure reducing valve 11 is connected to the pressure receiving portion 71c of the pressure compensating valve 7b via the signal lines 21 and 24, and connected to the pressure receiving portion 72c of the pressure compensating valve 7c via the signal lines 21 and 25. Then, the output pressure of the differential pressure reducing valve 11 is guided to the pressure receiving portions 71c and 72c as the target compensation differential pressure as described above.
[0029]
The pump tilt control mechanism 12 includes a horsepower control tilt actuator 12a that reduces the tilt of the hydraulic pump 2 when the discharge pressure of the hydraulic pump 2 is increased, and the discharge pressure of the hydraulic pump 2 is the highest of the plurality of actuators 3a, 3b, and 3c. An LS control valve 12b and an LS control tilt actuator 12c that perform load sensing control so as to be higher than the load pressure by a target differential pressure are provided.
[0030]
The LS control valve 12b includes a pressure receiving portion 12d located on the side that increases the pressure of the actuator 12c and reduces the tilt of the hydraulic pump 2, and a pressure receiving portion 12e located on the side that reduces the pressure of the actuator 12c and increases the tilt of the hydraulic pump 2. The pressure receiving portion 12d is supplied with the output pressure of the differential pressure reducing valve 11 (the pressure difference between the discharge pressure of the hydraulic pump 2 and the maximum load pressure of the actuators 3a, 3b, 3c), and the pressure receiving portion 12e is rotated with the engine. The output pressure of the differential pressure reducing valve 51 of the number detection circuit 13 is guided as a target differential pressure (target LS differential pressure) for load sensing control.
[0031]
  The pump tilt control mechanism 12 includes a pump tilt variable mechanism 90, and the actuators 12a and 12c operate the pump tilt variable mechanism 90 to control the tilt of the hydraulic pump 2. The pump tilt variable mechanism 90 is,oilA limiter 91 that restricts the tilt of the hydraulic pump 2 so as to set the minimum discharge flow rate of the pressure pump 2 to be equal to or higher than the maximum required flow rate is provided (described later).
[0032]
The engine speed detection circuit 13 includes a flow rate detection valve 50 and the above-described differential pressure reducing valve 51. The flow rate detection valve 50 includes a variable throttle portion 50a, and the throttle portion 50a is a discharge line of the pilot pump 30. 31. The discharge line 31 includes an upstream line 31a and a downstream line 31b of the flow rate detection valve 50, and a relief valve 32 that defines a source pressure as a pilot hydraulic pressure source is connected to the downstream line 31b. 31b is connected to a remote control valve (not shown) that generates a pilot pressure for switching the flow control valves 6a, 6b, 6c, for example.
[0033]
The flow rate detection valve 50 detects the flow rate of the pressure oil flowing through the discharge line 31 as a change in the differential pressure across the throttle 50a, and uses the differential pressure before and after the target LS differential pressure. Here, the flow rate of the pressure oil flowing through the discharge line 31 is the discharge flow rate of the pilot pump 30, and this discharge flow rate varies depending on the number of revolutions of the engine 1. This is to detect the rotational speed of the engine 1. For example, if the rotational speed of the engine 1 is decreased, the flow rate is decreased, and the differential pressure across the throttle portion 50a is decreased.
[0034]
Further, the throttle portion 50a is configured as a variable throttle portion whose opening area continuously changes, and the flow rate detection valve 50 further includes a pressure receiving portion 50b for opening direction operation, a pressure receiving portion 50c for throttle direction operation, and a spring 50d. The upstream pressure of the variable throttle portion 50a (pressure in the line 31a) is guided to the pressure receiving portion 50b, and the downstream pressure (pressure in the line 31b) of the variable throttle portion 50a is guided to the pressure receiving portion 50c, and the variable throttle portion 51a itself The opening area is changed depending on the differential pressure before and after. When the flow rate detection valve 50 is configured in this way and the differential pressure across the variable throttle 50a is used as the target LS differential pressure, the saturation phenomenon can be improved according to the engine speed, and the engine speed is set low. In addition, good fine operability can be obtained. This point is detailed in Japanese Patent Laid-Open No. 10-196604.
[0035]
The differential pressure reducing valve 51 is an engine speed detection valve that outputs the differential pressure before and after the variable restrictor 50a as an absolute pressure as a pressure that depends on the engine speed. Pressure receiving portions 51b and 51c located on the pressure reducing side, and the upstream pressure of the variable throttle portion 50a is guided to the pressure receiving portion 51a, and the downstream pressure of the variable throttle portion 50a and its own pressure are respectively received by the pressure receiving portions 51b and 51c. The output pressure is guided, and the differential pressure across the variable throttle 50a is output as an absolute pressure based on the pressure in the line 31b based on the balance of these pressures.
[0036]
The output port of the differential pressure reducing valve 51 is connected to the pressure receiving portion 12e of the LS control valve 12b via the signal line 53, and the output pressure of the differential pressure reducing valve 51 is guided to the pressure receiving portion 12e as the target LS differential pressure. As a result, the actuator speed can be set according to the engine speed. The output port of the differential pressure reducing valve 51 is connected to the pressure receiving portion 70c of the pressure compensating valve 7a via the signal lines 53 and 54, and the output pressure of the differential pressure reducing valve 51 is received as the target compensated differential pressure as described above. Guided to part 70c.
[0037]
  FIG. 2 is a PQ (pressure / flow rate) characteristic diagram of the hydraulic pump 2 by the pump tilt control mechanism 12. The horizontal axis is the discharge pressure of the hydraulic pump 2, the vertical axis is the discharge flow rate of the hydraulic pump 2, and the characteristic line H is a pump horsepower control diagram by the horsepower control tilt actuator 12a. QminIs oilThis is the minimum discharge flow rate of the pressure pump 2, QsmaxIs turningThe maximum required flow rate. The required maximum turning flow rate Qsmax is expressed by Qsmax = C · Amax√Pgr (C is a coefficient), where Amax is the maximum opening area of the meter-in throttle 61 or 62 of the turning flow control valve 6a. That is, generally, the required flow rate is a function of the opening area and its differential pressure before and after, and the differential pressure across the flow control valve 6a is determined by the pressure compensation valve 7a.ThatTarget compensation differenceTo pressureControlled to be equal,In the pressure receiving part 70c of the pressure compensation valve 7a,The target compensation differential pressureSince the output pressure of the differential pressure reducing valve 51 is derived as follows, the output pressure of the differential pressure reducing valve 51 is set to P gr IfQsmax = C · Amax√Pgr.
[0038]
As described above, the hydraulic pump 2 is limited by the limiter 91 so that the minimum discharge flow rate Qmin is equal to or greater than the maximum turning required flow rate Qsmax. For this reason, the pump LS control region by the LS control valve 12b and the LS control tilt actuator 12c is a shaded portion.
[0039]
Next, the operation of the present embodiment will be described in comparison with a comparative example. In the following description, the discharge pressure of the hydraulic pump 2 is Ps, the maximum load pressure of the actuators 3a, 3b, and 3c is PLmax, and the differential pressure (LS differential pressure) between the discharge pressure of the hydraulic pump 2 and the maximum load pressure is ΔPLS. The output pressure of the pressure reducing valve 11 is PLS, the output pressure of the differential pressure reducing valve 51 is Pgr, the target compensating differential pressure of the pressure compensating valve 7a is Pc1, the target compensating differential pressure of the pressure compensating valve 7b is Pc2, and the pressure compensating valve 7c The target compensation differential pressure is Pc3.
[0040]
FIG. 3 shows, as Comparative Example 1, not the output pressure of the differential pressure reducing valve 51 but the pressure compensating valves 7b and 7c of other sections in the pressure receiving portion 70c of the pressure compensating valve 7a of the swing section shown in FIG. In addition to guiding the output pressure of the differential pressure reducing valve 11, the pump tilt variable mechanism 90 of the pump tilt control mechanism 12X is not provided with the limiter 91, and the minimum discharge flow rate of the hydraulic pump 2 is set to a normal value. In this configuration, a differential pressure reducing valve 11 is added to the prior art described in JP-A-10-196604, and the discharge pressure Ps of the hydraulic pump 2 and the plurality of actuators 3a, 3b, 7c are added to the pressure compensating valves 7a, 7b, 7c. This corresponds to a configuration in which the differential pressure ΔPLS with respect to the maximum load pressure PLmax of 3c is guided as an absolute pressure (the output pressure PLS of the differential pressure reducing valve 11).
[0041]
In the comparative example 1 of FIG. 3, when a single actuator, for example, a single operation for driving the swing motor 3a is shifted to a composite operation for simultaneously driving a plurality of actuators, for example, the swing motor 3a and the boom cylinder 3b, the operability is improved. A decrease occurs.
[0042]
In a combined operation in which a plurality of actuators 3a and 3b are driven simultaneously, when the sum of the required flow rates for all the driven actuators exceeds the maximum flow rate that can be discharged by the hydraulic pump 2, the discharge flow rate of the hydraulic pump 2 is insufficient. A saturation condition occurs.
[0043]
In this case, the flow rate control valves 4a and 4b redistribute the flow rates to the actuators 3a and 3b. That is, the target compensation differential pressures Pc1 and Pc2 of the pressure compensation valves 7a and 7b related to the actuators 3a and 3b are given by a differential pressure (LS differential pressure) ΔPLS = Ps−PLmax between the discharge pressure Ps of the hydraulic pump 2 and the maximum load pressure PLmax. It is done. The LS differential pressure is supplied from the differential pressure reducing valve 11 as an absolute pressure PLS. In the saturation state, the LS differential pressure ΔPLS decreases according to the supply shortage ratio. For this reason, the target compensation differential pressures Pc1 and Pc2 of the pressure compensation valves 7a and 7b set by the differential pressures, which should be the same as the target LS differential pressure, are reduced by the supply shortage ratio. As a result, each actuator is supplied with a flow rate obtained by dividing the discharge flow rate of the hydraulic pump 2 into the required flow rate ratio. As a result, the overall actuator speed is reduced in accordance with the input ratio of the operation lever to each flow control valve 6a, 6b. This function makes it possible to maintain operability without stopping a specific actuator during a combined operation.
[0044]
However, in this combined operation, the speed of all the actuators is reduced uniformly, so that the operability may be reduced.
[0045]
For example, in the turning operation, the operator can feel the speed change, so that it is sensitive to the speed change, and it is necessary to have a configuration in which the speed change is suppressed as compared with other actuators. In addition, even in a suspended load operation or the like, if a speed change occurs during turning, the load will sway against the operator's intention.
[0046]
Further, as disclosed in Japanese Patent Laid-Open No. 2000-227103, a method of setting the turning priority spring in the pressure compensation valve 7a is also conceivable. However, in this case, since the target compensation differential pressure changes depending on the engine speed, it is impossible to maintain the priority at the same rate at all engine speeds.
[0047]
Therefore, the output pressure Pgr of the differential pressure reducing valve 51 which is a target value (target LS differential pressure) of LS control is led to the pressure receiving part 70c of the pressure compensating valve 7a of the swing section, and the target compensated differential pressure Pc1 is derived from the output pressure Pgr. A method of preventing the saturation information from being transmitted to the swing pressure compensating valve 7a by setting can be considered. FIG. 4 shows the configuration in that case as Comparative Example 2.
[0048]
However, in the above-described configuration, the pressure compensation valve 7a in the swing section and the pump tilt control mechanism 12 are both differential pressures across the meter-in restrictor 61 or 62 of the flow control valve 6a, which is the main spool of the swing section, during the single swing operation. Is controlled to the output pressure Pgr of the differential pressure reducing valve 51, and both interfere with each other.
[0049]
Hereinafter, the mechanism will be described.
[0050]
FIG. 5 is a PQ (pressure / flow rate) characteristic diagram of the hydraulic pump 2 similar to FIG. 2 by the pump tilt control mechanism 12X. Qmin0 is the minimum discharge flow rate of the normal hydraulic pump 2 when the limiter 91 is not provided. The minimum discharge flow rate Qmin0 of the hydraulic pump 2 is smaller than the maximum turning required flow rate Qsmax, and the maximum turning required flow rate Qsmax is included in the pump LS control region by the LS control valve 12b and the LS control tilting actuator 12c.
[0051]
In the single operation of turning, when the acceleration is started immediately after the start of turning and reaches the target speed and the acceleration is stopped, the differential pressure ΔPLS between the discharge pressure Ps of the hydraulic pump 2 and the maximum load pressure PLmax is reduced due to the decrease of the turning load pressure. It becomes larger than the output pressure Pgr of the differential pressure reducing valve 51 which is a control target value. As a result, the pump tilt control mechanism 12 controls the pump discharge pressure Ps of the pump supply oil passage 5 to decrease, so that the tilt of the hydraulic pump 2 becomes small and the discharge flow rate decreases. On the other hand, the pressure compensation valve 7a in the swing section moves in the closing direction in order to lower the differential pressure across the meter-in throttle 61 or 62 of the flow control valve 6a, and the pressure oil in the pump supply oil passage 5 is confined so that the pump discharge pressure. Ps is difficult to decrease. That is, while the LS control of the hydraulic pump 2 is controlled to reduce the discharge flow rate and lower the pump discharge pressure Ps, the pressure compensation valve 7a moves in the closing direction so that the discharge pressure Ps of the hydraulic pump 2 is less likely to decrease. Control. Here, the minimum discharge flow rate Qmax0 of the hydraulic pump 2 is a value smaller than the turning maximum required flow rate Qsmax. As a result, the discharge flow rate of the hydraulic pump 2 is extremely reduced by the LS control. Thus, the operation of the hydraulic pump 2 and the operation of the swing pressure compensation valve 7a interfere with each other, and the controllability of the discharge pressure of the hydraulic pump 2 is lowered.
[0052]
After that, when the discharge flow rate of the hydraulic pump 2 is consumed by the rotation of the swing motor 3a, the supply flow rate becomes insufficient because the discharge flow rate of the hydraulic pump 2 is extremely reduced. As a result, the differential pressure ΔPLS between the discharge pressure Ps of the hydraulic pump 2 and the maximum load pressure PLmax becomes lower than Pgr, which is the target value for the LS control. At this time, the pressure compensating valve 7a in the swing section is fully opened and does not function. At the same time, the pump tilt control mechanism 12 performs control so that the differential pressure ΔPLS between the pump discharge pressure Ps and the maximum load pressure PLmax becomes larger than Pgr which is the target value of LS control, that is, the direction in which the discharge pressure of the hydraulic pump 2 is increased. However, the tilt of the hydraulic pump 2 increases and the discharge flow rate increases. At that time, since the pressure compensation valve 7a of the swing section is fully opened, the discharge pressure Ps of the hydraulic pump 2 is difficult to increase, the control responsiveness is deteriorated, and the discharge flow rate of the hydraulic pump 2 is extremely increased by the LS control. .
[0053]
By repeating the above phenomenon, discontinuous acceleration occurs, and the operability is significantly reduced. This phenomenon is remarkably seen in turning with inertial load.
[0054]
In order to solve such a problem, in the present embodiment, as described above, the output pressure of the differential pressure reducing valve 51 is guided to the pressure receiving portion 70c of the pressure compensating valve 7a of the swing section, and the minimum discharge flow rate of the hydraulic pump 2 is obtained. Is set higher than the maximum required flow rate for turning.
[0055]
FIG. 6 is a diagram showing changes in the target compensation differential pressures Pc1, Pc2, and Pc3 of the pressure compensation valves 7a, 7b, and 7c in the present embodiment. In the figure, A is a single turning operation (non-saturation state), B is a turning combined operation (saturation state), both when the engine 1 is at the rated high speed, and C is a turning combined operation (saturation). State) when the engine 1 is at a low speed.
[0056]
The target compensation differential pressure Pc1 of the pressure compensation valve 7a in the swing section is the output pressure Pgr of the differential pressure reduction valve 51, which is the target value of LS control, and the target compensation differential pressure Pc2, of the other pressure compensation valves 7b, 7c, Pc3 is controlled to be the output pressure PLS of the differential pressure reducing valve 11 (absolute pressure of LS differential pressure ΔPLS).
[0057]
First, when the engine 1 is maintained at the rated high rotation speed, the output pressure Pgr of the differential pressure reducing valve 51 does not change, and the pressure compensating valve for the swing section during both the single swing operation and the combined operation (saturation state). The target compensation differential pressure Pc1 of 7a is constant (A and B in FIG. 6). For this reason, when the discharge flow rate of the hydraulic pump is sufficient during the single swing operation (A in FIG. 6), the output pressure PLS of the differential pressure reducing valve 11 is controlled to be equal to Pgr (PLS = Pgr). There is no fundamental difference in the control operation by the pressure compensation valves 7a, 7b, and 7c of the control valve 4 between the turning using Pgr as the target compensation differential pressure and the other actuators targeting PLS.
[0058]
In the combined operation including the turning (saturation state), the control valve 4 redistributes the flow rate to each actuator.
[0059]
First, the target compensation differential pressures Pc2 and Pc3 of the pressure compensation valves 7b and 7c other than the swing section decrease according to the ratio of insufficient supply of the pump discharge flow rate when the saturation state is reached. That is, the target compensation differential pressures Pc2 and Pc3 of the pressure compensation valves 7b and 7c are the differential pressure ΔPLS (= Ps−PLmax) between the discharge pressure Ps of the hydraulic pump 2 and the maximum load pressure PLmax, and the absolute pressure PLS by the differential pressure reducing valve 11. As given. In the saturation state, the LS differential pressure ΔPLS decreases according to the supply shortage ratio. For this reason, as shown in FIG. 6B, the target compensation differential pressures Pc2 and Pc3 of the pressure compensation valves 7b and 7c set by the differential pressure also decrease accordingly. As a result, the flow rate obtained by dividing the discharge flow rate of the hydraulic pump 2 into the respective required flow rate ratios is supplied to the actuators 3b and 3c except for the turning. As a result, the overall actuator speed is reduced in accordance with the input ratio of the operation lever to each flow control valve 6b, 6c.
[0060]
However, for turning, the target compensation differential pressure Pc1 of the pressure compensation valve 7a is set not by the LS differential pressure but by the output pressure Pgr of the differential pressure reducing valve 51, and this Pgr is the engine speed even in the saturation state. As long as it is at the rated high speed (B in FIG. 6), the saturation information is not transmitted to the pressure compensation valve 7a. As a result, the decrease in Pc1 in the saturation state is suppressed, and the pressure oil can be preferentially supplied to the swing motor 3a as compared with the other actuators 3b and 3c, and the swing motor 3a by the combined operation including the swing can be performed. It becomes possible to suppress the speed down.
[0061]
Further, when the engine speed is lowered to a low speed, the output pressure Pgr of the differential pressure reducing valve 51 decreases accordingly, so that the target compensation differential pressure Pc1 (= Pgr) of the pressure compensating valve 7a also decreases. As for the target compensation differential pressures Pc2 and Pc3 of the pressure compensation valves 7b and 7c other than the swing section, the differential pressure ΔPLS (= Ps−PLmax) between the discharge pressure Ps of the hydraulic pump 2 and the maximum load pressure PLmax when the engine 1 reaches a low speed. Is controlled so as to decrease in accordance with the output pressure Pgr of the differential pressure reducing valve 51, the target compensation differential pressures Pc2 and Pc3 also decrease. That is, as shown in FIG. 6C, in the combined turning operation (saturation state) when the engine speed is reduced, the target compensation differential pressure Pc1 of the pressure compensation valve 7a and the target compensation difference of the pressure compensation valves 7b and 7c. Both the pressures Pc2 and Pc3 decrease. For this reason, it does not happen that the turn becomes too fast, and appropriate turning priority can be maintained regardless of the setting of the engine speed.
[0062]
Next, the operation when shifting to the steady state from the start of the turning single operation will be described.
[0063]
At the time of turning start, the differential pressure ΔPLS between the discharge pressure Ps of the hydraulic pump 2 and the maximum load pressure PLmax becomes smaller than Pgr which is a target value for LS control. Therefore, the hydraulic pump 2 is controlled to increase the pump discharge pressure Ps of the pump supply oil passage 5, the pump tilt increases, and the discharge flow rate increases. At the same time, the pressure compensation valve 7a in the swing section is fully opened and does not function because the differential pressure across the meter-in throttle 61 or 62 (= ΔPLS) is smaller than the target compensation differential pressure Pc1 (= Pgr).
[0064]
When the turning speed reaches the target value and acceleration is lost, the differential pressure ΔPLS between the discharge pressure Ps of the hydraulic pump 2 and the maximum load pressure P1max becomes larger than Pgr, which is the target value for LS control. The hydraulic pump 2 is controlled so as to lower the pump discharge pressure Ps of the pump supply oil passage 5, so that the tilt of the hydraulic pump 2 is reduced and the discharge flow rate is reduced. At the same time, the pressure compensation valve 7a in the swing section operates in the closing direction, and controls the differential pressure across the meter-in throttle 61 or 62 to be the target compensation differential pressure Pc1 (= Pgr).
[0065]
  Here, the hydraulic pump 2 is driven by a limiter 91.MostThe small discharge flow rate Qmin is limited to be equal to or greater than the turning maximum required flow rate Qsmax. For this reason, when the discharge flow rate of the hydraulic pump 2 reaches the minimum discharge flow rate Qmin, it is maintained at Qmin thereafter, and the LS control of the hydraulic pump 2 does not operate. That is, the hydraulic pump 2 operates as if it is a fixed pump with a capacity Qmin. As a result, there is no interference between the LS control of the hydraulic pump 2 and the swing pressure compensation valve 7a.
[0066]
7 and 8 show the movement of the pump and the swing pressure compensating valve during the swing operation in the comparative example 2 shown in FIG. 4 and the present invention shown in FIG. In contrast to FIG. 1, the LS differential pressure ΔPLS (= Ps−PLmax) is summarized in tabular form when it becomes smaller than Pgr which is the target value of LS control and when it becomes larger.
[0067]
As described above, according to the present embodiment, in a hydraulic drive device equipped with an LS system, even when a saturation state of the pump discharge flow rate occurs, pressure oil is preferentially supplied to a specific actuator 3a to change its speed. The priority can be maintained in the same manner regardless of the setting of the engine speed, and excellent operability can be realized.
[0068]
Further, interference between the load sensing control of the hydraulic pump 2 and the operation of the pressure compensation valves 7a to 7c can be avoided, and the stability of the system can be maintained.
[0069]
In the present embodiment, the differential pressure between the discharge pressure of the hydraulic pump 2 and the maximum load pressure is derived as an absolute pressure by the output pressure of the differential pressure reducing valve 11, and the differential pressure across the variable throttle 50a of the engine detection circuit 13 is derived. Is derived as an absolute pressure by the output pressure of the differential pressure reducing valve 51, but the discharge pressure and the maximum load pressure of the hydraulic pump 2, the upstream pressure and the downstream pressure of the variable restrictor 50a are separately guided, respectively. Also good.
[0070]
【The invention's effect】
According to the present invention, in a hydraulic drive apparatus equipped with an LS system, even when a saturation state of a pump discharge flow rate occurs, pressure oil can be preferentially supplied to a specific actuator to suppress a speed change thereof, and the engine Regardless of the setting of the number of rotations, the priority can be maintained in the same manner, and excellent operability can be realized.
[0071]
Further, interference between the load sensing control of the hydraulic pump and the operation of the pressure compensation valve can be avoided, and the stability of the system can be maintained.
[Brief description of the drawings]
FIG. 1 is a diagram showing an overall configuration of a hydraulic drive apparatus according to an embodiment of the present invention.
FIG. 2 is a diagram showing a PQ characteristic of a hydraulic pump by a pump tilt control mechanism.
3 is a view similar to FIG. 1 showing a configuration based on the prior art as Comparative Example 1. FIG.
4 is a view similar to FIG. 1 showing a countermeasure example of Comparative Example 1 as Comparative Example 2. FIG.
5 is a graph showing PQ characteristics of a hydraulic pump by a pump tilt control mechanism in Comparative Examples 1 and 2. FIG.
FIG. 6 is a diagram showing a change in target compensation differential pressure in the hydraulic drive device shown in FIG. 1;
7 is a diagram showing, in a tabular form, movements of a pump and a swing pressure compensating valve during a swing operation in Comparative Example 2. FIG.
FIG. 8 is a table showing the movement of the pump and the swing pressure compensating valve during the swing operation of the present embodiment in a table format.
[Explanation of symbols]
1 engine
2 Main hydraulic pump
3a Actuator (swivel motor 9)
3b Actuator (boom cylinder)
3c Actuator (arm cylinder)
4 Control valve
4a, 4b, 4c Flow control valve
5 Supply oil passage
6a, 6b, 6c Flow control valve
7a, 7b, 7c Pressure compensation valve
8a, 8b, 8c, 8d Load line
9a, 9b Shuttle valve
10 signal lines
11 Differential pressure reducing valve
11a, 11b, 11c pressure receiving part
12 Pump tilt control mechanism
12a Horsepower control tilt actuator
12b LS control valve
12c LS control tilt actuator
13 Engine speed detection circuit
31 Discharge line
32 relief valve
50 Flow rate detection valve
50a Aperture part
51 Differential pressure reducing valve
51a, 51b, 51c pressure receiving part
53, 54 signal line
60a, 60b, 60c Load port
61, 62 Meter-in variable aperture
70a, 70b, 70c pressure receiving part
71a, 71b, 71c pressure receiving part
72a, 72b, 72c pressure receiving part
90 Pump tilt variable mechanism
91 Limiter

Claims

An engine, a variable displacement hydraulic pump driven by the engine, a plurality of actuators including a swing motor driven by pressure oil discharged from the hydraulic pump, and the hydraulic pump supplied to the plurality of actuators A plurality of flow control valves for controlling the flow of pressure oil, a plurality of pressure compensation valves for controlling the differential pressure across the plurality of flow control valves, respectively, and a discharge pressure of the hydraulic pump is a maximum load pressure of the plurality of actuators A pump control means for performing load sensing control so as to increase only by a target differential pressure, and an engine speed detection circuit for outputting a pressure depending on the engine speed so as to decrease as the engine speed decreases. direct the output pressure of the engine speed detecting circuit in the pump control means, the output of the engine speed detecting circuit Wherein by setting the target differential pressure of load sensing control, the rotation of the engine a target differential pressure of the load sensing control so that the target differential pressure of the load sensing control is reduced as the rotational speed of the engine is reduced by In the hydraulic drive device of a hydraulic excavator set as a variable value depending on the number,
Each target compensation differential pressure of the pressure compensation valve related to the actuator other than the swing motor is set by the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators, and the pressure compensation related to the swing motor An output pressure of the engine speed detection circuit is led to the valve, and a target compensation differential pressure is set by the output pressure of the engine speed detection circuit,
It said pump control means, the pump tilting variable mechanism comprising a limiter means minimum discharge flow rate to limit the tilting of the hydraulic pump so as not smaller than the maximum demanded flow rate of the flow control valve according to the swing motor before SL hydraulic pump A hydraulic drive device for a hydraulic excavator, comprising:

2. The hydraulic drive device for a hydraulic excavator according to claim 1, wherein the engine rotation speed detection circuit includes a first differential pressure reducing valve that outputs, as an absolute pressure, a pressure depending on the rotation speed of the engine, and other than the swing motor. A hydraulic drive device for a hydraulic excavator, wherein an output pressure of the first differential pressure reducing valve is guided to a pressure compensating valve related to the actuator of the actuator, and each target compensated differential pressure is set.

2. The hydraulic drive device for a hydraulic excavator according to claim 1, further comprising a second differential pressure reducing valve that outputs, as an absolute pressure, a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of actuators. A hydraulic excavator characterized in that an output pressure of the second differential pressure reducing valve is guided to a pressure compensating valve related to a motor, and a target compensated differential pressure of the pressure compensating valve is set by the output pressure of the second differential pressure reducing valve. Hydraulic drive device.