JP3705750B2

JP3705750B2 - Hydraulic drive

Info

Publication number: JP3705750B2
Application number: JP2001124180A
Authority: JP
Inventors: 智浜本; 康治岡崎; 行章長尾; 靖貴釣賀; 隆史金井; 純也川本
Original assignee: Nachi Fujikoshi Corp; Hitachi Construction Machinery Co Ltd
Current assignee: Nachi Fujikoshi Corp; Hitachi Construction Machinery Co Ltd
Priority date: 2001-04-23
Filing date: 2001-04-23
Publication date: 2005-10-12
Anticipated expiration: 2021-04-23
Also published as: JP2002323002A

Description

【０００１】
【発明の属する技術分野】
本発明は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方向切換弁の前後差圧をそれぞれ圧力補償弁により制御する油圧式ショベル等の油圧駆動装置に係わり、特に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力との差圧により設定し、かつロードセンシング制御の目標差圧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に関する。
【０００２】
【従来の技術】
出願人が特許出願中で未公開の特願2000-004074 号の図１に、図４で示すような油圧駆動装置が提案されている。図４では、エンジン1 と、このエンジン1 により駆動される可変容量型の油圧ポンプ10と、この油圧ポンプ10から吐出される圧油により駆動される複数のアクチュエータ 4a,4bと、油圧ポンプ10から複数のアクチュエータ 4a,4bに供給される圧油の流量をそれぞれ制御する複数の方向切換弁 20a,20bと、複数の方向切換弁の前後差圧をそれぞれ制御する複数の圧力補償弁 21a,21bと、油圧ポンプ10の吐出圧の上限を規制するメインリリーフ弁30とを備え、複数の圧力補償弁 21a,21bのそれぞれの目標補償差圧を差圧減圧弁34を介して油圧ポンプ10の吐出圧(Ps)と複数のアクチュエータ 20a,20bの最高負荷圧力(Plmax) との差圧に等しい２次圧力（Pc）により設定すると共に、油圧ポンプ10の吐出圧が複数のアクチュエータ 4a,4bの最高負荷圧よりロードセンシング目標差圧だけ高くなるよう前記ロードセンシング目標差圧と前記２次圧力とが対抗して導かれるポンプ傾転角制御弁12を含むロードセンシング制御するポンプ制御手段 12'を備え、特開平11-196604 号公報に開示する、ロードンシング制御の目標差圧 (PGR)をエンジン1 の回転数に依存する可変値として設定するため、固定容量型のパイロット油圧ポンプ11の吐出ライン9 に設けられた流量検出弁50の絞り部 50a前後の差圧を取り出す圧力発生弁51とを有し、かつロードセンシング制御機能向上のため、特開平10-205501 号公報に開示する、油圧ポンプ10から吐出される圧油ライン7 に接続され、ロードセンシング制御の目標差圧(PGR) と複数のアクチュエータ 20a,20bの最高負荷圧力(Plmax) がそれぞれ導かれる操作駆動部 31a,31dを有し、ロードセンシング制御の目標差圧(PGR) の変更に合わせて自身の目標差圧（ΔPun ）も変更されるようにした、可変アンロード弁31を有する、油圧駆動装置の油圧回路を示す。
【０００３】
図４では、同様にロードセンシング制御機能向上のため、特願2000-004074 号で提案する、油圧ポンプ10の吐出圧がメインリリーフ弁30の設定圧(Pr)まで上昇するとき、複数の圧力補償弁 21a,21bの目標補償差圧として油圧ポンプ10の吐出圧(Ps)と複数のアクチュエータ 20a,20bの最高負荷圧力(Plmax) との差圧（Pc）とは異なる補正値(Plmax')を設定する目標補償差圧補正手段が設けられており、この目標補償差圧補正手段は、最高負荷圧ライン35に固定絞り32を介して接続され、この最高負荷圧ライン35に検出される最高負荷圧力(Plmax) の上限をメインリリーフ弁30の設定圧力(Pr)より補正値(Plmax')分だけ低くする（設定圧力 Plmax'= Pr-PGR + α: αは PGRより小さい値）信号圧可変リリーフ弁33が設けられている。
【０００４】
【発明が解決しようとする課題】
図４で示すような油圧駆動装置の回路構成では、以下の課題があった。即ち、油圧式ショベル等の建設機械で、２つ以上のアクチュエータを同時に駆動させた状態から、一方のアクチュエータを停止した時に、それまで２つ以上のアクチュエータで消費していた油圧ポンプ10の吐出流量が、一方の側が停止したことによって、油圧ポンプ10の吐出流量に余剰流量が生じてしまうことから、油圧ポンプ10の吐出圧が急上昇し、油圧ポンプ10の吐出圧(Ps)とアクチュエータの最高負荷圧(Plmax) との差圧（Pc）である２次圧力を検出する差圧減圧弁で発生する２次圧力も急上昇し、急上昇した２次圧力は同時に駆動していて停止しなかった側の圧力補償弁に作用してその出口流量を増加させて停止しなかった側のアクチュエータに油圧ポンプ10の余剰流量を流入させて加速してしまう、という課題があった。即ち、同時に駆動させている２のアクチュエータをアームシリンダ4aとブームシリンダ4bとし、ブームシリンダ4bの操作を停止した場合、油圧ポンプ10の余剰流量がアームシリンダ4aに流入し、アームシリンダ4aを加速してしまい操作性の悪化を招くこととなっていた。
【０００５】
本来急上昇した２次圧力は、ロードセンシング制御するポンプ制御手段 12'ののポンプ傾転角制御弁12にも作用し、油圧ポンプ10の吐出流量を減少させ、余剰流量が出ないように制御する筈であるが、実際には油圧ポンプ10に応答遅れがあるため、急にポンプの吐出流量は減少しない。
【０００６】
本発明の課題は、油圧式ショベル等の建設機械で、２つ以上のアクチュエータを同時に駆動させた状態から、一方のアクチュエータを停止させた時においても、停止しなかった他方のアクチュエータにショックを発生させることがない回路構成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】
前述した課題解決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る圧油により駆動される複数のアクチュエータと、前記油圧ポンプから前記複数のアクチュエータに供給される圧油の流量をそれぞれ制御する複数の方向切換弁と、前記複数の方向切換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の上限を規制するメインリリーフ弁とを備え、前記複数の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力との差圧に等しい２次圧力により設定すると共に、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力よりロードセンシング目標差圧だけ高くなるよう前記ロードセンシング目標差圧と前記２次圧力とが対抗して導かれるポンプ傾転角制御弁を含むロードセンシング制御するポンプ制御手段を備え、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、かつ前記油圧ポンプから吐出される圧油ラインに接続され、前記ロードセンシング制御の目標差圧と前記複数のアクチュエータの前記最高負荷圧力がそれぞれ導かれる操作駆動部を有し、前記ロードセンシング制御の目標差圧の変更に合わせて自身の目標差圧も変更されるようにした、可変アンロード弁を有する油圧駆動装置において、
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力との前記差圧(Pc)を、前記差圧減圧弁から全ての前記圧力補償弁に、第１の絞りを有する差圧ラインを介して導くと共に、前記ポンプ傾転角制御弁へは前記差圧減圧弁からの前記２次圧力(Pc)を直接導いたことを特徴とする油圧駆動装置を提供することにより、上述した本発明の課題を解決した。
【０００８】
【発明の効果】
図４の油圧駆動装置の回路では、全ての圧力補償弁に対して、差圧減圧弁で得られるポンプの吐出圧と最高負荷圧力との差圧（Pc）に等しい２次圧力を作用させて、全てのアクチュエータに等しい比率でポンプの吐出流量を配分するアンチサチュレーション機能を得ていたのであるが、本発明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力との差圧に等しい２次圧力(Pc)を、差圧減圧弁から全ての圧力補償弁に、第１の絞りを介した差圧ラインを介して導く、ポンプ傾転角制御弁へは、差圧減圧弁からの前記差圧に等しい２次圧力(Pc)を直接導き、全ての前記圧力補償弁への２次圧力 (Pc')は第１の絞りの下流の差圧ラインから前記差圧(Pc)より第１の絞り抵抗分だけ低くした。かつ前記ポンプ傾転角制御弁へは前記差圧減圧弁からの前記２次圧力(Pc)を直接導いた。かかる構成により、差圧減圧弁からの２次圧力(Pc)と第１の絞りの下流の２次圧力 (Pc')とは静的にはほぼ等しい圧力となり、各圧力補償弁は従来通りの作用をしてアンチサチュレーション機能を有することになる。
ここで上記した２つ以上のアクチュエータを同時に駆動させた状態から、一方のアクチュエータを停止した場合について考えると、油圧ポンプの吐出圧が急上昇し、差圧減圧弁で発生する２次圧力(Pc)も急上昇するが、各圧力補償弁へ作用している２次圧力 (Pc')は、動的な変化に対して前記２次圧力(Pc)より第１の絞り抵抗分だけ低くした第１の絞りの下流側の圧力を導いていることから、第１の絞りの上流側の２次圧力(Pc)が急上昇しても、各圧力補償弁に導かれている２次圧力 (Pc')は、急上昇を緩和された圧力 (Pc')になって伝達される結果、各圧力補償弁により制御されているアクチュエータへの供給流量は急激に増大することがなくなった。さらに油圧ポンプのポンプ傾転角制御弁へ導かれる２次圧力(Pc)は第１の絞りの上流側の圧力であることから、２次圧力(Pc)の急上昇はそのままポンプ傾転角制御弁に通じるので、油圧ポンプの流量制御の応答性を損なうことがなく、良好な制御性を得ることができる。
【０００９】
好ましくは、前記第１の絞りの下流の全ての圧力補償弁との間の前記差圧ラインを第２の絞りを介してタンクと連通させることにより、第１の絞りの下流側の２次圧力 (Pc')をタンクへと通じさせている。
ここで第１の絞りと第２の絞りとの絞り径による開口面積の関係は、
第１の絞り＞＞第２の絞り
として、差圧減圧弁からの２次圧力(Pc)と第１の絞りの下流の２次圧力 (Pc')とに大きな差異が生じないようにしている。かかる構成により、第１の絞りの下流側の２次圧力 (Pc')はタンクへと通じていることから、２次圧力(Pc)の急上昇は第１の絞りと第２の絞りを介してタンクと連通し、下流側の２次圧力 (Pc')には直接伝達しないことになり、従って、上述したように、各圧力補償弁に導かれている２次圧力 (Pc')は、急上昇を緩和された圧力 (Pc')になって伝達される結果、各圧力補償弁により制御されているアクチュエータへの供給流量は急激に増大することがなくなり、さらに２次圧力(Pc)の急上昇はそのままポンプ傾転角制御弁に通じるので、油圧ポンプの流量制御の応答性を損なうことがなく、良好な制御性を得ることができる。
【００１０】
さらに好ましくは、各複数の前記圧力補償弁の少なくとも１つははそれぞれ自己負荷圧が上昇すると自身の出口吐出量が減少することにより、自己負荷圧が急上昇したアクチュエータへの流量を下げてポンプの必要吐出量を下げて他方のアクチュエータが停止した際のポンプ圧力の急上昇を防ぎ、それにより差圧減圧弁で発生する２次圧力(Pc)の急上昇を防ぎ、２次圧力の急上昇に伴うポンプ吐出量の急減少を防ぎ、油圧駆動装置の複数のアクチュエータを同時に操作する複合操作時の複合操作性を向上させることができる。
【００１１】
より好ましくは、前記油圧ポンプの吐出圧が前記メインリリーフ弁の設定圧まで上昇するとき、前記複数の圧力補償弁の目標補償差圧として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力との差圧とは異なる補正値を設定する目標補償差圧補正手段を設け、前記目標補償差圧補正手段は、前記最高負荷圧力を検出する最高負荷圧ラインに固定絞りを介して接続され、この最高負荷圧ラインに検出される最高負荷圧力の上限を前記メインリリーフ弁の設定圧力よりも前記補正値分だけ低くする信号圧可変リリーフ弁を有するようにして、複数のアクチュエータを同時に操作する複合操作時にどれか１つのアクチュエータの負荷圧がメインリリーフ弁の設定圧力に達しても、圧力補償弁が閉弁せず、かつ他のアクチュエータが増速せず、油圧駆動装置の複合操作性をより向上させることができる。
【００１２】
【発明の実施の形態】
図１は本発明の第１の実施の形態である油圧駆動装置の油圧回路を示す。本発明の第１の実施の形態の油圧回路では、図４の回路で説明したと同様に、エンジン1 と、このエンジン1 により駆動される可変容量型の油圧ポンプ10と、この油圧ポンプ10から吐出される圧油により駆動される複数のアクチュエータ 4a,4b（うち、２個のみ示す）と、油圧ポンプ10から複数のアクチュエータ 4a,4bに供給される圧油の流量をそれぞれ制御する複数の方向切換弁 20a,20b（うち、２個のみ示す）と、複数の方向切換弁の前後差圧をそれぞれ制御する複数の圧力補償弁 21a,21b（うち、２個のみ示す）と、油圧ポンプ10の吐出圧が複数のアクチュエータ 4a,4bの最高負荷圧力(Plmax) 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5 及びポンプ傾転角制御弁12と、油圧ポンプ10の吐出圧力の上限を規制するメインリリーフ弁30と、を有する。
【００１３】
そして特開平11-196604 号公報に開示する、油圧ポンプ10の吐出圧が複数のアクチュエータ 4a,4bの最高負荷圧力(Plmax) よりロードセンシング目標差圧だけ高くなるよう前記ロードセンシング目標差圧と前記２次圧力とが対抗して導かれるポンプ傾転角制御弁12を含むロードセンシング制御するポンプ制御手段12' を備え、ロードセンシング制御の目標差圧 (PGR)発生回路5 を有する。即ち、ロードセンシング制御の目標差圧 (PGR)をエンジン1 の回転数に依存する可変値として設定するため、固定容量型のパイロット油圧ポンプ11の吐出ライン9 に設けられた流量検出弁50の絞り部 50a前後の差圧 (PGR)として取り出す圧力発生弁51とを有する。かつロードセンシング制御機能向上のため、特開平10-205501 号公報に開示する、油圧ポンプ10から吐出される圧油ライン7 に接続され、ロードセンシング制御の目標差圧(PGR) と複数のアクチュエータ 20a,20bの最高負荷圧力(Plmax) がそれぞれ導かれる操作駆動部 31a,31dを有し、ロードセンシング制御の目標差圧(PGR) の変更に合わせて自身の目標差圧（ΔPun ）も変更されるようにした、可変アンロード弁31を有し、エンジン低回転時の微小流量制御特性を向上させている。これらロードセンシング制御の目標差圧 (PGR)発生回路5 、可変アンロード弁31、の各構成・作用の詳細は上記公報に記載済であり、重複した説明はしない。
【００１４】
本発明では、油圧ポンプ10の吐出圧と複数のアクチュエータ 4a,4bの最高負荷圧力(Plmax) との差圧に等しい２次圧力（Pc）を、差圧減圧弁34から全ての圧力補償弁 21a,21bに、第１の絞り61を有する差圧ライン2 を介して導くと共に、ポンプ傾転角制御弁12へは、差圧減圧弁34からの差圧(Pc)を直接導き、全ての圧力補償弁 21a,21bへの２次圧力 (Pc')は第１の絞り61の下流の差圧ライン2 から前記差圧(Pc)より第１の絞り61の絞り抵抗分だけ低くした。さらに油圧ポンプのポンプ傾転角制御弁12へ２次圧力(Pc)は第１の絞り61の上流側の圧力を直接導いた。
【００１５】
かかる構成により、差圧減圧弁34からの２次圧力(Pc)と第１の絞り61の下流の２次圧力 (Pc')とは静的にはほぼ等しい圧力となり、各圧力補償弁 21a,21bは従来通りの作用をしてアンチサチュレーション機能を有することになる。
ここで上記した２つ以上のアクチュエータ 4a,4bを同時に駆動させた状態から、一方のアクチュエータを停止した場合について考えると、油圧ポンプ10の吐出圧が急上昇し、差圧減圧弁34で発生する２次圧力(Pc)も急上昇するが、各圧力補償弁 21a,21bへ作用している２次圧力 (Pc')は、動的な変化に対して前記２次圧力(Pc)より第１の絞り抵抗分だけ低くした第１の絞り61の下流側の圧力を導いていることから、第１の絞り61の上流側の２次圧力(Pc)が急上昇しても、各圧力補償弁 21a,21bに導かれている２次圧力 (Pc')は、急上昇を緩和された圧力 (Pc')になって伝達される結果、各圧力補償弁 21a,21bにより制御されているアクチュエータ 4a,4bへの供給流量は急激に増大することがなくなった。さらに油圧ポンプのポンプ傾転角制御弁12へ導かれる２次圧力(Pc)は第１の絞り61の上流側の圧力であることから、２次圧力(Pc)の急上昇はそのままポンプ傾転角制御弁12に通じるので、油圧ポンプ10の流量制御の応答性を損なうことがなく、良好な制御性を得ることができる。
【００１６】
図２は本発明の第２の実施の形態である油圧駆動装置の油圧回路を示す。図２では、第１の絞り61の下流の全ての圧力補償弁 21a,21bとの間の差圧ライン2 を、第２の絞り62を介してタンク3 と連通させた。
ここで、第１の絞り61と第２の絞り62との絞り径による開口面積の関係は、
第１の絞り＞＞第２の絞り
として、差圧減圧弁34からの２次圧力(Pc)と第１の絞り61の下流の２次圧力 (Pc')とに大きな差異が生じないようにしている。かかる構成により、第１の絞り61の下流側の２次圧力 (Pc')はタンク3 へと通じていることから、２次圧力(Pc)の急上昇は第１の絞り61と第２の絞り62を介してタンク3 と連通し、下流側の２次圧力 (Pc')には直接伝達しないことになり、従って、各圧力補償弁 21a,21bに導かれている２次圧力 (Pc')は、急上昇を緩和された圧力 (Pc')になって伝達される結果、各圧力補償弁 21a,21bにより制御されているアクチュエータ 4a,4bへの供給流量は急激に増大することがなくなった。さらに２次圧力(Pc)の急上昇はそのままポンプ傾転角制御弁12に通じるので、油圧ポンプ10の流量制御の応答性を損なうことがなく、良好な制御性を得ることができる。
【００１７】
図３は図１及び図２の圧力補償弁 21a,21bの少なくとも１つの好ましい実施の形態の概略構成を示す圧力補償弁 21A,21Bの断面図であり、特開平10-089304 号公報特に図３に詳細構成・作用を示しており、重複した説明を省略する。圧力補償弁 21A,21Bの室 121の受圧面A1には出口 105の圧力Pzが導かれ、室 113の受圧面A2には差圧減圧弁34で得られる２次圧力Pcが導かれ、室 119の受圧面A3には自己負荷圧Plが導かれる。ここまでは、図１及び図２の圧力補償弁 21a,21bと同じであるが、図３の圧力補償弁 21A,21Bでは、スプール 117の直径d1は、スプール 112の室 119内の小径部の直径d3より大にされ、
d3＜d1 Ａ３＜Ａ１、Ａ３＝ｋ・Ａ１（但しｋ＜１）にされている。
スプール 112は左右から下記力を受けて均衡する。
(A3・Pl) ＋ (A2・Pc) ＝ (A1・Pz)
これにより、各圧力補償弁 21A,21Bが補償する各方向切換弁 20a,20b前後の差圧ΔＰは、次式で表すことができる。
ΔＰ＝ [(k・A2)/A3] ・Pc - (1-k)・Pl
このため、複数の圧力補償弁 21a,21bのそれぞれ自己負荷圧Plが上昇すると、自己負荷圧が急上昇したアクチュエータへの方向切換弁前後の差圧ΔＰが減少して、圧力補償弁及び方向切換弁の各出口吐出量が減少することにより、自己負荷圧が急上昇したアクチュエータへの流量を下げてポンプの必要吐出量を下げて他方のアクチュエータが停止した際のポンプ圧力の急上昇を防ぎ、それにより差圧減圧弁で発生する２次圧力(Pc)の急上昇を防ぎ、２次圧力の急上昇に伴うポンプ吐出量の急減少を防ぎ、油圧駆動装置の複数のアクチュエータを同時に操作する複合操作時の複合操作性を向上させることができる。
【００１８】
さらに図１及び図２に示す本発明の実施の形態では、ロードセンシング制御機能向上のため、特願2000-004074 号で提案する、油圧ポンプ10の吐出圧がメインリリーフ弁30の設定圧(Pr)まで上昇するとき、複数の圧力補償弁 21a,21bの目標補償差圧として油圧ポンプ11の吐出圧(Ps)と複数のアクチュエータ 20a,20bの最高負荷圧力(Plmax) との差圧（Pc）とは異なる補正値(Plmax')を設定する目標補償差圧補正手段が設けられており、この目標補償差圧補正手段は、最高負荷圧ライン35に設けられ、この最高負荷圧ライン35に検出される最高負荷圧力(Plmax)の上限をメインリリーフ弁30の設定圧力(Pr)より補正値(Plmax')分だけ低くする（設定圧力 Plmax'= Pr-PGR + α: αはPGR より小さい値）信号圧可変リリーフ弁33が設けるようにして、複数のアクチュエータを同時に操作する複合操作時にどれか１つのアクチュエータの負荷圧力がメインリリーフ弁の設定圧力に達しても、圧力補償弁が閉弁せず、かつ他のアクチュエータが増速せず、油圧駆動装置の複合操作性をより向上させることができる。
【図面の簡単な説明】
【図１】本発明の第１の実施の形態である油圧駆動装置の油圧回路を示す。
【図２】本発明の第２の実施の形態である油圧駆動装置の油圧回路を示す。
【図３】図１及び図２の圧力補償弁 21a,21bの少なくとも１つの好ましい実施の形態の概略構成を示す圧力補償弁 21A,21Bの断面図である。
【図４】出願人が特許出願中で未公開の特願2000-004074 号の図１に示す油圧駆動装置の油圧回路を示す。
【符号の説明】
１・・エンジン 2 ・・差圧ライン
3 ・・タンク 4a,4b ・・アクチュエータ
7 ・・圧油ライン 10・・可変容量型の油圧ポンプ
11・・パイロット油圧ポンプ 12・・ポンプ傾転角制御弁
20a,20b ・・方向切換弁 21a,21b,21A,21B ・・圧力補償弁
30・・メインリリーフ弁 31・・可変アンロード弁
33・・信号圧可変リリーフ弁 35・・最高負荷圧ライン
61・・第１の絞り 62・・第２の絞り[0001]
BACKGROUND OF THE INVENTION
The present invention is a hydraulic excavator or the like that performs load sensing control so that the discharge pressure of a hydraulic pump is higher than a maximum load pressure of a plurality of actuators by a target differential pressure, and controls the differential pressure across a plurality of directional control valves by a pressure compensation valve The target compensation differential pressure of each pressure compensation 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 is engine. The present invention relates to a hydraulic drive device that is set as a variable value that depends on the number of rotations.
[0002]
[Prior art]
A hydraulic drive device as shown in FIG. 4 is proposed in FIG. 1 of Japanese Patent Application No. 2000-004074, for which the applicant has not applied for a patent. In FIG. 4, the engine 1, a variable displacement hydraulic pump 10 driven by the engine 1, a plurality of actuators 4 a and 4 b driven by pressure oil discharged from the hydraulic pump 10, and the hydraulic pump 10 A plurality of directional control valves 20a, 20b that respectively control the flow rates of pressure oil supplied to the plurality of actuators 4a, 4b, and a plurality of pressure compensation valves 21a, 21b that respectively control the differential pressure across the plurality of directional control valves; And a main relief valve 30 that regulates the upper limit of the discharge pressure of the hydraulic pump 10, and the target compensation differential pressure of each of the plurality of pressure compensation valves 21a and 21b is supplied to the discharge pressure of the hydraulic pump 10 via the differential pressure reducing valve 34. The secondary pressure (Pc) is equal to the differential pressure between (Ps) and the maximum load pressure (Plmax) of the multiple actuators 20a, 20b, and the discharge pressure of the hydraulic pump 10 is the maximum load of the multiple actuators 4a, 4b. Load sensing target differential pressure than pressure And a pump control means 12 ′ for load sensing control including a pump tilt angle control valve 12 in which the load sensing target differential pressure and the secondary pressure are opposed to each other so as to be higher, and disclosed in JP-A-11-196604 The flow rate detection valve 50 provided in the discharge line 9 of the fixed displacement type pilot hydraulic pump 11 is set in order to set the target differential pressure (PGR) of the road-and-singing control disclosed in FIG. Pressure generating valve 51 for taking out the differential pressure before and after the throttle part 50a, and for improving the load sensing control function, disclosed in Japanese Patent Laid-Open No. 10-205501, is a pressure oil line 7 discharged from the hydraulic pump 10. To the load sensing control target differential pressure (PGR) and the maximum load pressure (Plmax) of each of the actuators 20a and 20b. PGR) 1 shows a hydraulic circuit of a hydraulic drive apparatus having a variable unload valve 31 in which its own target differential pressure (ΔPun) is also changed in accordance with the change of
[0003]
In FIG. 4, in order to improve the load sensing control function as well, when the discharge pressure of the hydraulic pump 10 rises to the set pressure (Pr) of the main relief valve 30 proposed in Japanese Patent Application No. 2000-004074, a plurality of pressure compensations are performed. A correction value (Plmax ') that is different from the differential pressure (Pc) between the discharge pressure (Ps) of the hydraulic pump 10 and the maximum load pressure (Plmax) of the multiple actuators 20a, 20b as the target compensated differential pressure of the valves 21a, 21b The target compensation differential pressure correction means to be set is provided, and this target compensation differential pressure correction means is connected to the maximum load pressure line 35 via the fixed throttle 32 and is detected by the maximum load pressure line 35. Lower the upper limit of the pressure (Plmax) by the correction value (Plmax ') from the set pressure (Pr) of the main relief valve 30 (set pressure Plmax' = Pr-PGR + α: α is a value smaller than PGR) and variable signal pressure A relief valve 33 is provided.
[0004]
[Problems to be solved by the invention]
The circuit configuration of the hydraulic drive device as shown in FIG. 4 has the following problems. That is, when two or more actuators are driven simultaneously by a construction machine such as a hydraulic excavator, when one actuator is stopped, the discharge flow rate of the hydraulic pump 10 that has been consumed by the two or more actuators until then. However, when one side stops, an excessive flow rate is generated in the discharge flow rate of the hydraulic pump 10, so that the discharge pressure of the hydraulic pump 10 suddenly rises, and the discharge pressure (Ps) of the hydraulic pump 10 and the maximum load of the actuator The secondary pressure generated by the differential pressure reducing valve that detects the secondary pressure that is the differential pressure (Pc) with the pressure (Plmax) also rises rapidly. The suddenly raised secondary pressure is driven simultaneously and has not stopped. There has been a problem that the surplus flow rate of the hydraulic pump 10 is caused to flow into the actuator that has not stopped by acting on the pressure compensation valve to increase its outlet flow rate, thereby accelerating. That is, when the two actuators driven simultaneously are the arm cylinder 4a and the boom cylinder 4b and the operation of the boom cylinder 4b is stopped, the surplus flow rate of the hydraulic pump 10 flows into the arm cylinder 4a and accelerates the arm cylinder 4a. As a result, the operability deteriorates.
[0005]
The secondary pressure that has suddenly increased also acts on the pump tilt angle control valve 12 of the pump control means 12 'that performs load sensing control, thereby reducing the discharge flow rate of the hydraulic pump 10 and controlling so that no excessive flow rate is generated. Although the hydraulic pump 10 actually has a response delay, the pump discharge flow rate does not suddenly decrease.
[0006]
The subject of the present invention is that a construction machine such as a hydraulic excavator generates a shock to the other actuator that has not stopped even when one actuator is stopped from the state where two or more actuators are driven simultaneously. An object of the present invention is to provide a hydraulic drive device having a circuit configuration that is not allowed to occur.
[0007]
[Means for Solving the Problems]
In order to solve the problems described above, the present invention includes an engine, a variable displacement hydraulic pump driven by the engine, a plurality of actuators driven by pressure oil discharged from the hydraulic pump, and the hydraulic pump. A plurality of directional control valves for controlling the flow rates of pressure oil supplied to the plurality of actuators; a plurality of pressure compensating valves for controlling differential pressures before and after the plurality of directional switching valves; and a discharge pressure of the hydraulic pump A main relief valve that regulates the upper limit of the pressure, and a target compensation differential pressure of each of the plurality of pressure compensation valves is equal to a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of actuators. The discharge pressure of the hydraulic pump is set higher by the load sensing target differential pressure than the maximum load pressure of the plurality of actuators. And a pump control means for performing load sensing control including a pump tilt angle control valve through which the load sensing target differential pressure and the secondary pressure are opposed to each other, and the target differential pressure of the load sensing control is determined by rotating the engine. An operation drive unit that is set as a variable value that depends on the number and that is connected to a pressure oil line that is discharged from the hydraulic pump, and that guides the target differential pressure of the load sensing control and the maximum load pressure of the plurality of actuators, respectively. In the hydraulic drive apparatus having a variable unload valve, the target differential pressure of the load sensing control is changed in accordance with the change of the target differential pressure of the load sensing control.
The differential pressure (Pc) between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators is transferred from the differential pressure reducing valve to all the pressure compensating valves via a differential pressure line having a first throttle. And providing the hydraulic pressure drive device characterized in that the secondary pressure (Pc) from the differential pressure reducing valve is directly guided to the pump tilt angle control valve. Solved the problem.
[0008]
【The invention's effect】
In the circuit of the hydraulic drive apparatus of FIG. 4, a secondary pressure equal to the differential pressure (Pc) between the pump discharge pressure and the maximum load pressure obtained by the differential pressure reducing valve is applied to all the pressure compensating valves. The anti-saturation function that distributes the pump discharge flow rate at an equal ratio to all the actuators was obtained. In the present invention, the secondary pressure equal to the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of a plurality of actuators is obtained. The pressure (Pc) is led from the differential pressure reducing valve to all the pressure compensating valves via the differential pressure line via the first throttle, and the difference from the differential pressure reducing valve is supplied to the pump tilt angle control valve. A secondary pressure (Pc) equal to the pressure is directly derived, and the secondary pressures (Pc ′) to all the pressure compensating valves are first from the differential pressure line (Pc) from the differential pressure line downstream of the first throttle. The diaphragm resistance is lowered. The secondary pressure (Pc) from the differential pressure reducing valve was directly introduced to the pump tilt angle control valve. With this configuration, the secondary pressure (Pc) from the differential pressure reducing valve and the secondary pressure (Pc ') downstream of the first throttle are statically substantially equal, and each pressure compensating valve is the same as the conventional one. It acts to have an anti-saturation function.
Here, considering the case where one of the actuators is stopped from the state in which two or more actuators are driven simultaneously, the discharge pressure of the hydraulic pump rises rapidly, and the secondary pressure (Pc) generated by the differential pressure reducing valve. However, the secondary pressure (Pc ′) acting on each pressure compensation valve is less than the secondary pressure (Pc) by the first throttle resistance against the dynamic change. Since the pressure on the downstream side of the throttle is guided, even if the secondary pressure (Pc) on the upstream side of the first throttle suddenly rises, the secondary pressure (Pc ') led to each pressure compensation valve is As a result, the sudden increase is transmitted as a moderated pressure (Pc ′). As a result, the supply flow rate to the actuator controlled by each pressure compensation valve does not increase rapidly. Further, since the secondary pressure (Pc) guided to the pump tilt angle control valve of the hydraulic pump is the pressure upstream of the first throttle, the sudden rise in the secondary pressure (Pc) remains as it is. Therefore, good controllability can be obtained without impairing the responsiveness of the flow control of the hydraulic pump.
[0009]
Preferably, the secondary pressure downstream of the first throttle is communicated with the tank via the second throttle in the differential pressure line between all the pressure compensating valves downstream of the first throttle. (Pc ') is connected to the tank.
Here, the relationship of the opening area depending on the aperture diameter of the first aperture and the second aperture is:
As the first throttle >> second throttle, the secondary pressure (Pc) from the differential pressure reducing valve and the secondary pressure (Pc ') downstream of the first throttle are prevented from greatly differing. . With this configuration, since the secondary pressure (Pc ′) downstream of the first throttle is connected to the tank, the sudden rise in the secondary pressure (Pc) is caused through the first throttle and the second throttle. It communicates with the tank and does not transmit directly to the secondary pressure (Pc ') on the downstream side. Therefore, as described above, the secondary pressure (Pc') led to each pressure compensation valve increases rapidly. As a result, the supply flow rate to the actuator controlled by each pressure compensation valve does not increase rapidly, and the secondary pressure (Pc) increases rapidly. Since it passes directly to the pump tilt angle control valve, the responsiveness of the flow control of the hydraulic pump is not impaired, and good controllability can be obtained.
[0010]
More preferably, at least one of the plurality of pressure compensating valves has its own outlet discharge amount reduced when the self-load pressure increases, thereby decreasing the flow rate to the actuator where the self-load pressure suddenly increased, thereby reducing the pump flow rate. Reduces the required discharge amount and prevents a sudden increase in pump pressure when the other actuator stops, thereby preventing a sudden increase in secondary pressure (Pc) generated by the differential pressure reducing valve. Pump discharge due to a sudden increase in secondary pressure It is possible to prevent a sudden decrease in the amount and improve the composite operability at the time of the composite operation in which a plurality of actuators of the hydraulic drive device are operated simultaneously.
[0011]
More preferably, when the discharge pressure of the hydraulic pump rises to a set pressure of the main relief valve, the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators as target compensation differential pressures of the plurality of pressure compensation valves And a target compensation differential pressure correction means for setting a correction value different from the differential pressure, and the target compensation differential pressure correction means is connected to a maximum load pressure line for detecting the maximum load pressure via a fixed throttle, A composite that operates a plurality of actuators simultaneously by having a signal pressure variable relief valve that lowers the upper limit of the maximum load pressure detected in the maximum load pressure line by the correction value from the set pressure of the main relief valve. Even if the load pressure of one of the actuators reaches the set pressure of the main relief valve during operation, the pressure compensation valve does not close and other actuators Not accelerated, it is possible to further improve the operability in the combined operation of the hydraulic drive system.
[0012]
DETAILED DESCRIPTION OF THE INVENTION
FIG. 1 shows a hydraulic circuit of a hydraulic drive apparatus according to a first embodiment of the present invention. In the hydraulic circuit according to the first embodiment of the present invention, as described with reference to the circuit of FIG. 4, the engine 1, the variable displacement hydraulic pump 10 driven by the engine 1, and the hydraulic pump 10 Plural actuators 4a, 4b (only two are shown) driven by the discharged pressure oil, and plural directions for controlling the flow rate of the pressure oil supplied from the hydraulic pump 10 to the plurality of actuators 4a, 4b, respectively Of the switching valves 20a and 20b (of which only two are shown), a plurality of pressure compensating valves 21a and 21b (of which only two are shown) for controlling the differential pressure across the plurality of directional switching valves, respectively, Pump control means 5 that performs load sensing control so that the discharge pressure is higher than the maximum load pressure (Plmax) of the multiple actuators 4a, 4b by the target differential pressure, the pump tilt angle control valve 12, and the upper limit of the discharge pressure of the hydraulic pump 10 Regulating the main lily It has a valve 30, a.
[0013]
And, as disclosed in JP-A-11-196604, the load sensing target differential pressure and the discharge pressure of the hydraulic pump 10 are set so as to be higher than the maximum load pressure (Plmax) of the plurality of actuators 4a, 4b by the load sensing target differential pressure. A pump control means 12 ′ for load sensing control including a pump tilt angle control valve 12 which is guided against the secondary pressure is provided, and a target differential pressure (PGR) generation circuit 5 for load sensing control is provided. That is, in order to set the target differential pressure (PGR) of load sensing control as a variable value depending on the rotation speed of the engine 1, the throttle of the flow rate detection valve 50 provided in the discharge line 9 of the fixed displacement pilot hydraulic pump 11 is set. And a pressure generating valve 51 to be taken out as a differential pressure (PGR) before and after the part 50a. In order to improve the load sensing control function, a target differential pressure (PGR) for load sensing control and a plurality of actuators 20a are connected to the pressure oil line 7 discharged from the hydraulic pump 10 disclosed in JP-A-10-205501. , 20b, the maximum load pressure (Plmax) is guided, respectively, and the target differential pressure (ΔPun) is changed according to the change of the target differential pressure (PGR) of load sensing control. Thus, the variable unload valve 31 is provided to improve the minute flow rate control characteristic at the time of low engine rotation. Details of the configuration and action of the target differential pressure (PGR) generation circuit 5 and the variable unload valve 31 of the load sensing control have already been described in the above publication, and will not be redundantly described.
[0014]
In the present invention, the secondary pressure (Pc) equal to the differential pressure between the discharge pressure of the hydraulic pump 10 and the maximum load pressure (Plmax) of the plurality of actuators 4a, 4b is supplied from the differential pressure reducing valve 34 to all the pressure compensating valves 21a. , 21b through the differential pressure line 2 having the first restrictor 61, and the differential pressure (Pc) from the differential pressure reducing valve 34 is directly guided to the pump tilt angle control valve 12 so that all pressures are The secondary pressure (Pc ′) to the compensation valves 21a and 21b was made lower than the differential pressure (Pc) by the throttle resistance of the first throttle 61 from the differential pressure line 2 downstream of the first throttle 61. Further, the secondary pressure (Pc) directly led to the upstream pressure of the first throttle 61 to the pump tilt angle control valve 12 of the hydraulic pump.
[0015]
With this configuration, the secondary pressure (Pc) from the differential pressure reducing valve 34 and the secondary pressure (Pc ′) downstream of the first throttle 61 are statically substantially equal pressures, and each pressure compensating valve 21a, 21b has an anti-saturation function by acting as before.
Considering a case where one of the actuators 4a and 4b is simultaneously driven and then stopping one of the actuators, the discharge pressure of the hydraulic pump 10 rises rapidly and is generated by the differential pressure reducing valve 34. Although the secondary pressure (Pc) also rises sharply, the secondary pressure (Pc ') acting on each pressure compensating valve 21a, 21b is the first throttle than the secondary pressure (Pc) with respect to dynamic changes. Since the pressure on the downstream side of the first throttle 61 reduced by the resistance is introduced, even if the secondary pressure (Pc) on the upstream side of the first throttle 61 suddenly rises, each pressure compensating valve 21a, 21b The secondary pressure (Pc ′) led to the pressure is transferred to the pressure compensation valve 21a, 21b controlled by the pressure compensating valves 21a, 21b as a result of being transmitted as the pressure (Pc ′) with the sudden rise reduced. The supply flow rate no longer increases rapidly. Furthermore, since the secondary pressure (Pc) guided to the pump tilt angle control valve 12 of the hydraulic pump is the pressure upstream of the first throttle 61, the sudden rise in the secondary pressure (Pc) remains as it is. Since the control valve 12 is communicated, the controllability of the flow rate control of the hydraulic pump 10 is not impaired and good controllability can be obtained.
[0016]
FIG. 2 shows a hydraulic circuit of a hydraulic drive apparatus according to the second embodiment of the present invention. In FIG. 2, the differential pressure line 2 between all the pressure compensating valves 21 a and 21 b downstream of the first throttle 61 is communicated with the tank 3 via the second throttle 62.
Here, the relationship of the opening area by the diameter of the first diaphragm 61 and the second diaphragm 62 is as follows:
First throttle >> As a second throttle, make sure that there is no significant difference between the secondary pressure (Pc) from the differential pressure reducing valve 34 and the secondary pressure (Pc ') downstream of the first throttle 61. ing. With this configuration, since the secondary pressure (Pc ′) downstream of the first throttle 61 communicates with the tank 3, the sudden increase in the secondary pressure (Pc) causes the first throttle 61 and the second throttle to increase. It communicates with the tank 3 via 62 and is not directly transmitted to the downstream secondary pressure (Pc '). Therefore, the secondary pressure (Pc') guided to each pressure compensating valve 21a, 21b As a result, the sudden increase is transmitted as a moderated pressure (Pc ′). As a result, the supply flow rate to the actuators 4a and 4b controlled by the pressure compensating valves 21a and 21b does not increase rapidly. Further, since the rapid increase in the secondary pressure (Pc) is directly passed to the pump tilt angle control valve 12, the responsiveness of the flow rate control of the hydraulic pump 10 is not impaired and good controllability can be obtained.
[0017]
FIG. 3 is a cross-sectional view of pressure compensating valves 21A and 21B showing a schematic configuration of at least one preferred embodiment of the pressure compensating valves 21a and 21b of FIGS. 1 and 2, and Japanese Patent Laying-Open No. 10-089304, particularly FIG. The detailed configuration / operation is shown in FIG. The pressure Pz of the outlet 105 is guided to the pressure receiving surface A1 of the chamber 121 of the pressure compensation valves 21A and 21B, and the secondary pressure Pc obtained by the differential pressure reducing valve 34 is guided to the pressure receiving surface A2 of the chamber 113, and the chamber 119 The self-load pressure Pl is guided to the pressure receiving surface A3. The pressure compensation valves 21a and 21b shown in FIGS. 1 and 2 are the same up to this point, but in the pressure compensation valves 21A and 21B shown in FIG. 3, the diameter d1 of the spool 117 is smaller than that of the small diameter portion in the chamber 119 of the spool 112. Is made larger than the diameter d3,
d3 <d1 A3 <A1, A3 = k · A1 (where k <1).
The spool 112 is balanced by receiving the following force from the left and right.
(A3 ・ Pl) + (A2 ・ Pc) = (A1 ・ Pz)
Thereby, the differential pressure ΔP before and after each directional control valve 20a, 20b compensated by each pressure compensating valve 21A, 21B can be expressed by the following equation.
ΔP = [(k ・ A2) / A3] ・ Pc-(1-k) ・ Pl
Therefore, when the self-load pressure Pl of each of the plurality of pressure compensation valves 21a and 21b increases, the differential pressure ΔP before and after the direction switching valve to the actuator where the self-load pressure suddenly increases decreases, and the pressure compensation valve and the direction switching valve By reducing the discharge amount at each outlet, the flow rate to the actuator whose self-load pressure has increased rapidly is reduced, the required discharge amount of the pump is reduced, and the sudden increase in pump pressure when the other actuator stops is prevented. Prevents sudden increase in secondary pressure (Pc) generated by the pressure reducing valve, prevents sudden decrease in pump discharge amount due to sudden increase in secondary pressure, and operates multiple actuators at the same time to operate multiple actuators of a hydraulic drive unit Can be improved.
[0018]
Further, in the embodiment of the present invention shown in FIGS. 1 and 2, the discharge pressure of the hydraulic pump 10 proposed in Japanese Patent Application No. 2000-004074 is set to the set pressure (Pr) of the main relief valve 30 in order to improve the load sensing control function. ), The differential pressure (Pc) between the discharge pressure (Ps) of the hydraulic pump 11 and the maximum load pressure (Plmax) of the multiple actuators 20a, 20b as the target compensation differential pressure of the multiple pressure compensation valves 21a, 21b Target compensation differential pressure correction means for setting a different correction value (Plmax ') is provided. This target compensation differential pressure correction means is provided in the maximum load pressure line 35 and is detected in this maximum load pressure line 35. Lower the upper limit of the maximum load pressure (Plmax) to be set by the correction value (Plmax ') from the set pressure (Pr) of the main relief valve 30 (set pressure Plmax' = Pr-PGR + α: α is smaller than PGR) ) Operate multiple actuators at the same time as the signal pressure variable relief valve 33 is provided. Even if the load pressure of one of the actuators reaches the set pressure of the main relief valve during combined operation, the pressure compensation valve does not close and the speed of other actuators does not increase, thus improving the combined operability of the hydraulic drive device. It can be improved further.
[Brief description of the drawings]
FIG. 1 shows a hydraulic circuit of a hydraulic drive apparatus according to a first embodiment of the present invention.
FIG. 2 shows a hydraulic circuit of a hydraulic drive device according to a second embodiment of the present invention.
3 is a cross-sectional view of pressure compensation valves 21A and 21B showing a schematic configuration of at least one preferred embodiment of the pressure compensation valves 21a and 21b of FIGS. 1 and 2. FIG.
FIG. 4 shows a hydraulic circuit of the hydraulic drive device shown in FIG. 1 of Japanese Patent Application No. 2000-004074 that has not been filed by the applicant.
[Explanation of symbols]
1 ・・ Engine 2 ・・ Differential pressure line
3 .. Tanks 4a, 4b
7 ·· Pressure oil line 10 · · Variable displacement hydraulic pump
11 ・・ Pilot hydraulic pump 12 ・・ Pump tilt angle control valve
20a, 20b ・・ Direction switching valve 21a, 21b, 21A, 21B ・・ Pressure compensation valve
30 ・・ Main relief valve 31 ・・ Variable unloading valve
33 ・・ Signal pressure variable relief valve 35 ・・ Maximum load pressure line
61 .. First aperture 62 .. Second aperture

Claims

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る圧油により駆動される複数のアクチュエータと、前記油圧ポンプから前記複数のアクチュエータに供給される圧油の流量をそれぞれ制御する複数の方向切換弁と、前記複数の方向切換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の上限を規制するメインリリーフ弁とを備え、前記複数の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力との差圧に等しい２次圧力により設定すると共に、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力よりロードセンシング目標差圧だけ高くなるよう前記ロードセンシング目標差圧と前記２次圧力とが対抗して導かれるポンプ傾転角制御弁を含むロードセンシング制御するポンプ制御手段を備え、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、かつ前記油圧ポンプから吐出される圧油ラインに接続され、前記ロードセンシング制御の目標差圧と前記複数のアクチュエータの前記最高負荷圧力がそれぞれ導かれる操作駆動部を有し、前記ロードセンシング制御の目標差圧の変更に合わせて自身の目標差圧も変更されるようにした、可変アンロード弁を有する油圧駆動装置において、
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力との前記差圧(Pc)を、前記差圧減圧弁から全ての前記圧力補償弁に、第１の絞りを有する差圧ラインを介して導くと共に、前記ポンプ傾転角制御弁へは前記差圧減圧弁からの前記２次圧力(Pc)を直接導いたことを特徴とする油圧駆動装置。An engine, a variable displacement hydraulic pump driven by the engine, a plurality of actuators driven by pressure oil discharged from the hydraulic pump, and a pressure oil supplied from the hydraulic pump to the plurality of actuators. A plurality of directional control valves that respectively control the flow rate, a plurality of pressure compensation valves that control the differential pressure across the plurality of directional control valves, and a main relief valve that regulates the upper limit of the discharge pressure of the hydraulic pump. The target compensation differential pressure of each of the plurality of pressure compensation valves is set by a secondary pressure equal to the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators, and the discharge of the hydraulic pump The load sensing target differential pressure and the previous pressure so that the pressure is higher than the maximum load pressure of the plurality of actuators by the load sensing target differential pressure. Pump control means for load sensing control including a pump tilt angle control valve that is guided against the secondary pressure is provided, and the target differential pressure of the load sensing control is set as a variable value depending on the engine speed. And an operation drive unit that is connected to a pressure oil line that is discharged from the hydraulic pump and that guides the target differential pressure of the load sensing control and the maximum load pressure of the plurality of actuators, respectively. In the hydraulic drive device having a variable unload valve, the target differential pressure is changed in accordance with the change of the target differential pressure.
The differential pressure (Pc) between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators is transferred from the differential pressure reducing valve to all the pressure compensating valves via a differential pressure line having a first throttle. And the secondary pressure (Pc) from the differential pressure reducing valve is directly guided to the pump tilt angle control valve.

前記第１の絞りの下流の全ての圧力補償弁との間の前記差圧ラインを第２の絞りを介してタンクと連通させたことを特徴とする請求項１記載の油圧駆動装置。2. The hydraulic drive apparatus according to claim 1, wherein the differential pressure line between all pressure compensating valves downstream of the first throttle is communicated with the tank via the second throttle.

各複数の前記圧力補償弁の少なくとも１つはそれぞれ自己負荷圧が上昇すると自身の出口吐出量が減少するようにしたことを特徴とする請求項１又は請求項２記載の油圧駆動装置。3. The hydraulic drive apparatus according to claim 1, wherein at least one of the plurality of pressure compensation valves has its outlet discharge amount decreased when the self-load pressure increases.

前記油圧ポンプの吐出圧が前記メインリリーフ弁の設定圧まで上昇するとき、前記複数の圧力補償弁の目標補償差圧として、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力との差圧とは異なる補正値を設定する目標補償差圧補正手段を設け、前記目標補償差圧補正手段は、前記最高負荷圧力を検出する最高負荷圧ラインに固定絞りを介して接続され、この最高負荷圧ラインに検出される最高負荷圧力の上限を前記メインリリーフ弁の設定圧力よりも前記補正値分だけ低くする信号圧可変リリーフ弁を有することを特徴とする請求項１乃至請求項３のいずれかに記載の油圧駆動装置。When the discharge pressure of the hydraulic pump rises to the set pressure of the main relief valve, the difference between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators is used as the target compensation differential pressure of the plurality of pressure compensation valves. Target compensation differential pressure correction means for setting a correction value different from the pressure is provided, and the target compensation differential pressure correction means is connected to a maximum load pressure line for detecting the maximum load pressure via a fixed throttle, and this maximum load 4. The signal pressure variable relief valve according to claim 1, further comprising a signal pressure variable relief valve that lowers an upper limit of the maximum load pressure detected in the pressure line by the correction value from a set pressure of the main relief valve. The hydraulic drive device described in 1.