JP3853123B2 - Hydraulic drive - Google Patents

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JP3853123B2
JP3853123B2 JP34382399A JP34382399A JP3853123B2 JP 3853123 B2 JP3853123 B2 JP 3853123B2 JP 34382399 A JP34382399 A JP 34382399A JP 34382399 A JP34382399 A JP 34382399A JP 3853123 B2 JP3853123 B2 JP 3853123B2
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pressure
differential pressure
compensation
valve
turning
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JP2000227103A (en
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靖貴 釣賀
隆史 金井
純也 川本
賢一郎 中谷
究 高橋
智 浜本
康治 岡崎
行章 長尾
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Nachi Fujikoshi Corp
Hitachi Construction Machinery Co Ltd
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Nachi Fujikoshi Corp
Hitachi Construction Machinery Co Ltd
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Description

【０００１】
【発明の属する技術分野】
本発明は、油圧ショベル等、旋回制御系を含む建設機械の油圧駆動装置に係わり、特に旋回モータを含む複数のアクチュエータにそれぞれの方向切換弁を介して油圧ポンプからの圧油を供給する際に、油圧ポンプの吐出流量をロードセンシングシステムにより制御しかつ方向切換弁の前後差圧をそれぞれの圧力補償弁により制御する油圧駆動装置に関する。
【０００２】
【従来の技術】
油圧ポンプの吐出流量をロードセンシングシステム（以下、適宜ＬＳシステムという）により制御する油圧駆動装置として、特開昭６０−１１７０６号公報に記載のものがある。また、旋回制御系を含む建設機械の油圧駆動装置でＬＳシステムを備えかつ旋回制御系の独立性と操作性を実現するものとして、特開平１０−３７９０７号公報に記載のものがある。更に、旋回制御系を含む建設機械のオープンセンタタイプの油圧駆動装置で旋回制御系の独立性を実現するものとして、実機搭載の３ポンプシステムがある。更に、油圧ポンプの吐出流量をＬＳシステムにより制御する油圧駆動装置で圧力補償弁に負荷依存特性を持たせたものとして、特開平１０−８９３０４号公報に記載のものがある。
特開昭６０−１１７０６号公報に記載の油圧駆動装置は、複数の圧力補償弁のそれぞれに、油圧ポンプの吐出圧力と複数のアクチュエータの最高負荷圧との差圧を目標補償差圧として設定する手段を設けたものであり、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方向切換弁のの要求する流量に満たないサチュレーション状態になると、このサチュレーション状態により油圧ポンプの吐出圧力と最高負荷圧の差圧が低くなることにより、圧力補償弁のそれぞれの目標補償差圧が小さくなり、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できる。
【０００３】
特開平１０−３７９０７号公報に記載の油圧駆動装置及び実機搭載の３ポンプシステムは、いずれも、旋回モータを含む旋回セクションに関して、独立した油圧ポンプを用いたオープンセンタタイプの独立した回路により他のアクチュエータと別回路を構成し、旋回制御系の独立性と操作性を確保したものである。
【０００４】
特開平１０−８９３０４号公報に記載の油圧駆動装置は、複数の圧力補償弁のそれぞれについて、圧力補償弁の油圧室のうち、方向切換弁の入側圧力が導かれる閉じ方向作用の油圧室の受圧面積を、方向切換弁の出側圧力が導かれる開け方向作用の油圧室の受圧面積よりも大きくすることにより、各アクチュエータの負荷圧の増加に対して圧力補償弁の目標補償差圧を小さくし（圧力補償弁を絞り）、アクチュエータへの供給流量を減らす負荷依存特性を持たせたものであり、これにより低負荷側、高負荷側共操作性が良く、ハンチングを生じず、安定して動作し得るようになる。
【０００５】
【発明が解決しようとする課題】
しかしながら、上記従来の油圧駆動装置は、旋回制御系に関して次のような問題がある。
【０００６】
特開昭６０−１１７０６号公報：下記問題点▲１▼▲２▼
特開平１０−８９３０４号公報：下記問題点▲２▼▲３▼
特開平１０−３７９０７号公報：下記問題点▲４▼
実機搭載のオープンセンタタイプの３ポンプシステム：下記問題点▲４▼
▲１▼旋回単独起動時の操作性のギクシャク感
▲２▼旋回単独動作から旋回複合動作への移行時又はその逆の場合の旋回速度変化
▲３▼旋回複合起動時の旋回速度の極端な低下
▲４▼別回路を設けることによるコスト・スペースの増加及び回路構成の複雑化
【０００７】
（１）特開昭６０−１１７０６号公報
特開昭６０−１１７０６号公報に記載のＬＳシステムを備えた油圧駆動装置では、これを旋回制御系に用いた場合、旋回制御系は慣性負荷を伴うため、油圧ポンプのロードセンシング制御（以下、適宜ＬＳ制御という）と圧力補償弁の流量補償機能とのバランスが取り難くなる。これは、次の理由により、旋回加速時から定常回転へ移行する段階での旋回駆動圧力の制御に際して、圧力補償弁の応答性と油圧ポンプのＬＳ制御の応答性との間でバランスが取り難いことが挙げられる。
【０００８】
(1)旋回起動・加速時は、一定流量を保持するため、ポンプＬＳ制御は旋回起動圧に応じて油圧ポンプの吐出圧力を高く制御する。
【０００９】
(2)圧力補償弁は方向切換弁の絞り要素前後の差圧を一定に保持するため、負荷圧の上昇により低下する傾向にある通過流量を増やす方向に動作している。
【００１０】
(3)旋回が定常速度に達すると旋回駆動圧が下がるため、ポンプＬＳ制御は起動・加速時ほど油圧ポンプの吐出圧力を高く制御する必要がなく、油圧ポンプの吐出圧力を下げる方向に動作する。
【００１１】
(4)圧力補償弁は、旋回駆動圧の低下により、増加する傾向にある通過流量を減らす方向に動作する。
【００１２】
上記(1)〜(4)の移行が急峻なため、旋回操作性はギクシャクとしたものになる（上記▲１▼）。
【００１３】
また、上記のように複合動作時に、油圧ポンプの吐出流量が複数の方向切換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ると、このサチュレーション状態に応じて圧力補償弁のそれぞれの目標補償差圧が小さくなり、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。この機能により、複合動作時にもそれぞれのアクチュエータは、スピードダウンするものの、その動作を目的とした割合で動作するため、操作感を損なわない。
【００１４】
しかし、このスピードダウンは、旋回動作に関しても同様に発生し、旋回を含む複合動作時に旋回速度は他のアクチュエータと同じくスピードダウンする。このスピードダウンは、旋回複合動作から旋回単独動作に移行する場合、又はその逆の場合には旋回速度の変化を生じ、オペレータに違和感を与える（上記▲２▼）。
【００１５】
（２）特開平１０−８９３０４号公報
特開平１０−８９３０４号公報に記載の油圧駆動装置は、圧力補償弁に負荷依存特性を持たせたため、旋回単独起動時、旋回モータの高圧の負荷圧に応じて圧力補償弁の目標補償差圧が低下し、定常状態に移行すると旋回モータの低下した負荷圧に応じて圧力補償弁の目標補償差圧も元に戻り、これにより旋回操作性のギクシャク感なく旋回を起動できる。しかし、旋回複合動作時に油圧ポンプの吐出流量がサチュレーション状態になると、油圧ポンプの吐出流量をそれぞれの方向切換弁が要求する流量の比に再分配することは、特開昭６０−１１７０６号公報に記載の油圧駆動装置と同じであり、旋回複合動作から旋回単独動作に移行する場合、又はその逆の場合には旋回速度の変化を生じ、オペレータに違和感を与える（上記▲２▼）。
【００１６】
また、圧力補償弁に負荷依存特性を持たせているため、旋回複合起動時、旋回セクションの圧力補償弁は、油圧ポンプの吐出流量の状態に応じて圧力補償弁の目標補償差圧が小さくなるだけでなく、旋回モータの負荷圧がリリーフ圧まで上昇する負荷依存特性によっても目標補償差圧が低下し、この目標補償差圧の低下は定常状態に移行するまで持続する。この結果、旋回複合起動時の旋回速度が他のアクチュエータに比べて極端に低下し、旋回複合起動の旋回操作性が損なわれる（上記▲３▼）。
【００１７】
（３）特開平１０−３７９０７号公報に記載の油圧駆動装置や実機搭載のオープンセンタタイプの３ポンプシステム
特開平１０−３７９０７号公報に記載の油圧駆動装置では、旋回制御系をオープンセンタタイプの別回路で構成することにより、旋回操作性をＬＳシステムにおいて確保している。また、実機搭載のオープンセンタタイプの３ポンプシステムでも、旋回制御系はオープンセンタタイプの別回路であり、旋回操作性を確保してる。
【００１８】
即ち、オープンセンタタイプの場合、旋回起動時、駆動圧が上昇すると、センタバイパス油路を経てタンクに還流する流量が増えるため、旋回セクションの方向切換弁の絞りを通過する圧油の流量が減少する。このため、旋回モータに供給される圧油の流量は起動・加速時に制限される。旋回速度が定常速度に達すると、駆動圧は起動時ほど高くないため、流量の制限はなくなり、旋回セクションの方向切換弁の絞りの開口相当の流量が旋回モータに供給される。これによりＬＳ制御のような旋回単独起動時の操作性のギクシャク感を生じることなく、スムーズに旋回起動が行える。
【００１９】
また、上記▲２▼の問題はＬＳシステムのみによらず、オープンセンタタイプのシステムでも発生するが、特開平１０−３７９０７号公報に記載の油圧駆動装置や実機搭載のオープンセンタタイプの３ポンプシステムでは、旋回制御系をオープンセンタタイプの別回路で構成することにより、旋回制御系の独立性を実現し、旋回速度変化は生じない。
【００２０】
しかし、特開平１０−３７９０７号公報に記載の油圧駆動装置や実機搭載のオープンセンタタイプの３ポンプシステムでは、旋回制御系を、他のアクチュエータのシステムとは別回路で並列に構成しなくてはならなず、その分コスト高となりかつ設置スペースも大となると共に、旋回制御系用の油圧ポンプを別に設けなくてはならず、特に特開平１０−３７９０７号公報のシステムでは、並列に配置されるＬＳシステムとのパワーバランスをとるため、信号経路が必要となり、回路構成が複雑となる（上記▲４▼）。
【００２１】
本発明の目的は、旋回制御系を含む油圧駆動装置において、旋回単独、複合のいずれの起動時にも、旋回操作性のギクシャク感がなく加速して定常状態に移行でき、しかも旋回単独動作から旋回複合動作への移行時又はその逆の場合の旋回速度変化が抑えられ、かつ複合の起動時に他のアクチュエータに比べ旋回速度が極端に遅くならず、優れた旋回操作性と旋回独立性を確保できると共に、別回路を設けることによるコスト・スペースの増加や回路構成の複雑化の問題を生じない油圧駆動装置を提供することである。
【００２２】
【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、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力が前記複数のアクチュエータの最高負荷圧より所定値だけ高くなるようポンプ吐出流量を制御するロードセンシング制御のポンプ制御手段とを備えた油圧駆動装置において、前記複数の圧力補償弁のうち、前記旋回モータに係わる旋回セクション以外の圧力補償弁に設けられ、前記油圧ポンプの吐出圧力と前記複数のアクチュエータの最高負荷圧との差圧を目標補償差圧として設定する第１手段と、前記旋回セクションの圧力補償弁に設けられ、その目標補償差圧を設定する第２手段と、前記複数の圧力補償弁のうち、少なくとも前記旋回セクションの圧力補償弁に設けられ、前記旋回モータの負荷圧が上昇すると、前記第２手段で設定された目標補償差圧を小さくし、旋回セクションの圧力補償弁に負荷依存特性を持たせる第３手段と、前記旋回セクションの圧力補償弁に設けられ、前記第２手段で設定され、前記第３手段で補正される目標補償差圧の下限が、前記油圧ポンプの吐出圧力と前記複数のアクチュエータの最高負荷圧との差圧が低下したときに前記第１手段により設定される目標補償差圧よりも小さくならないように、前記第２手段で設定され、前記第３手段で補正される目標補償差圧の下限を設定する第４手段とを備えるものとする。
【００２３】
以上のように構成した本発明においては、旋回セクションの圧力補償弁に第３手段を設け負荷依存特性を持たせることにより、旋回起動時に旋回モータの負荷圧の変化に応じて旋回セクションの圧力補償弁は流量を微調整し、旋回モータはスムーズに加速して定常状態に移行するものとなる。
【００２４】
また、旋回セクションの圧力補償弁の目標補償差圧を設定する第２手段は、第１手段と同じように、油圧ポンプの吐出圧力と複数のアクチュエータの最高負荷圧との差圧を目標補償差圧として設定する手段であってもよく、この場合は、上記のように第４手段を設けることにより、この第４手段が第２手段で設定された目標補償差圧自体の低下と第３手段で与えられた負荷依存特性による目標補償差圧の低下の両方に対して下限設定手段として機能するものとなる（下記（２）参照）。これにより油圧ポンプの吐出流量がサチュレーション状態になり旋回セクションの圧力補償弁の目標補償差圧が低下しようとするとき、或いは旋回モータの負荷圧が高圧になり旋回セクションの圧力補償弁の目標補償差圧が負荷依存特性により低下しようとするとき、或いはそれらが同時に起こるとき、第４手段はその目標補償差圧の低下を制限し、旋回モータに優先的に圧油が供給されるものとなる。その結果、旋回単独動作から旋回複合動作への移行時又はその逆の場合の旋回速度変化が抑えられ、かつ複合の起動時に他のアクチュエータに比べ旋回速度が極端に遅くならず、優れた旋回操作性と旋回独立性を確保できる。
【００２５】
旋回セクションの圧力補償弁の目標補償差圧を設定する第２手段は、油圧ポンプの吐出圧力と複数のアクチュエータの最高負荷圧との差圧により変化しない値を目標補償差圧として設定する手段であってもよく、この場合は、第４手段は、第３手段で与えられた負荷依存特性による目標補償差圧の低下に対して下限設定手段として機能するものとなる（下記（３）参照）。これにより油圧ポンプの吐出流量がサチュレーション状態になっても、旋回セクションの圧力補償弁の目標補償差圧は低下せず、かつ旋回モータの負荷圧が高圧になり旋回セクションの圧力補償弁の目標補償差圧が負荷依存特性により低下しようとするとき、第４手段はその目標補償差圧の低下を制限し、サチュレーション或いは負荷依存特性による目標補償差圧の低下が単独或いは同時のいずれで起こっても、旋回モータに優先的に圧油が供給されるものとなる。その結果、旋回単独動作から旋回複合動作への移行時又はその逆の場合の旋回速度変化が抑えられ、かつ複合の起動時に他のアクチュエータに比べ旋回速度が極端に遅くならず、優れた旋回操作性と旋回独立性を確保できる。
【００２６】
更に、別回路を設けることなく上記の機能を達成するので、コスト・スペースの増加や回路構成の複雑化の問題も生じない。
【００２７】
（２）上記（１）において、好ましくは、前記第２手段は、前記第１手段と同様、前記油圧ポンプの吐出圧力と前記複数のアクチュエータの最高負荷圧との差圧を前記目標補償差圧として設定する手段であり、前記第４手段は、前記第２手段で設定された目標補償差圧自体の低下と前記第３手段で与えられた負荷依存特性による目標補償差圧の低下の両方に対して下限設定手段として機能する。
【００２８】
これにより上記（１）で述べたように、油圧ポンプの吐出流量がサチュレーション状態になり旋回セクションの圧力補償弁の目標補償差圧が低下しようとするとき、或いは旋回モータの負荷圧が高圧になり旋回セクションの圧力補償弁の目標補償差圧が負荷依存特性により低下しようとするとき、或いはそれらが同時に起こるとき、第４手段はその目標補償差圧の低下を制限し、旋回モータに優先的に圧油が供給されるものとなり、優れた旋回操作性と旋回独立性を確保できる。
【００２９】
（３）また、上記（１）において、前記第２手段は、前記油圧ポンプの吐出圧力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変化しない値を前記目標補償差圧として設定する手段であってもよく、この場合、前記第４手段は、前記第３手段で与えられた負荷依存特性による目標補償差圧の低下に対して下限設定手段として機能する。
【００３０】
これにより上記（１）で述べたように、油圧ポンプの吐出流量がサチュレーション状態になっても、旋回セクションの圧力補償弁の目標補償差圧は低下せず、かつ旋回モータの負荷圧が高圧になり旋回セクションの圧力補償弁の目標補償差圧が負荷依存特性により低下しようとするときは、第４手段はその目標補償差圧の低下を制限し、サチュレーション或いは負荷依存特性による目標補償差圧の低下が単独或いは同時のいずれで起こっても、旋回モータに優先的に圧油が供給されるものとなり、優れた旋回操作性と旋回独立性を確保できる。
【００３１】
（４）更に、上記（１）〜（３）において、好ましくは、前記第４手段は、前記第２手段で設定され、前記第３手段で補正される目標補償差圧が所定値に達すると、前記旋回セクションの圧力補償弁のスプールに開け方向の付勢力を付与する付勢手段である。
【００３２】
これにより第４手段は、付勢手段が付与する付勢力相当の値以下に旋回セクションの圧力補償弁の目標補償差圧を低下させず、目標補償差圧の下限を設定するものとなる。
【００３３】
（５）上記（４）において、好ましくは、前記付勢手段は、前記第２手段で設定され、前記第３手段で補正される目標補償差圧が所定値に達すると、前記旋回セクションの圧力補償弁のスプールに作用し、このスプールを開け方向に付勢する下限設定バネである。
【００３４】
これにより付勢手段は、旋回セクションの圧力補償弁の目標補償差圧が所定値に達すると、旋回セクションの圧力補償弁のスプールに開け方向の付勢力を付与するものとなる。
【００３５】
（６）また、上記（１）及び（２）において、好ましくは、前記第４手段は、前記第２手段で設定され、前記第３手段で補正される目標補償差圧に常時補助的な値を付加する付勢手段であり、前記旋回セクションの方向切換弁は、そのメータイン可変絞りの開口面積が、前記付勢手段で付加される補助的な値の目標補償圧相当分だけ、旋回セクション以外の方向切換弁の開口面積より小さくなるように構成されている。
【００３６】
これにより第４手段は、付勢手段で付加する補助的な値の分、旋回セクションの圧力補償弁の目標補償差圧の低下を制限し、目標補償差圧の下限を設定するものとなる。
【００３７】
（７）上記（６）において、好ましくは、前記付勢手段は、前記旋回セクションの圧力補償弁のスプールの開け方向に常時作用する旋回優先バネである。
【００３８】
これにより付勢手段は、旋回セクションの圧力補償弁の目標補償差圧に常時補助的な値を付加するものとなる。
【００３９】
【発明の実施の形態】
以下、本発明の実施形態を図面を用いて説明する。
【００４０】
図１は本発明の第１の実施形態による油圧駆動装置を示すものであり、油圧ポンプ１と、この油圧ポンプ１から吐出される圧油により駆動される旋回モータ２を含む複数のアクチュエータ２〜６と、油圧ポンプ１から複数のアクチュエータ２〜６に供給される圧油の流量をそれぞれ制御するクローズドセンタタイプの複数の方向切換弁７〜１１と、複数の方向切換弁７〜１１の前後差圧をそれぞれ制御する複数の圧力補償弁１２〜１６と、方向切換弁７〜１１と圧力補償弁１２〜１６との間に配置され、圧油の逆流を防止するロードチェック弁１７ａ〜１７ｅと、油圧ポンプ１の吐出圧力が複数のアクチュエータ２〜６の最高負荷圧より所定値だけ高くなるようポンプ吐出流量を制御するロードセンシング制御のポンプ制御装置１８とを備えている。旋回モータ２のアクチュエータラインにはオーバロードリリーフ弁６０ａ，６０ｂが設けられている。他のアクチュエータ３〜６にも同様なオーバロードリリーフ弁が設けられているが、図示は省略する。
【００４１】
複数の方向切換弁７〜１１には自己負荷圧の検出ライン２０〜２４が設けられ、これら検出ライン２０〜２４で検出された負荷圧のうちの最高負荷圧が信号ライン２５〜２９、シャトル弁３０〜３３及び信号ライン３４〜３６を介して検出され、信号ライン３７に導出される。
【００４２】
ポンプ制御装置１８は、油圧ポンプ１の容量可変部材である斜板１ａに連結された傾転制御アクチュエータ４０と、このアクチュエータ４０の油圧室４０ａと油圧ポンプ１の吐出油路１ｂ及びタンク１９との接続を切換制御するロードセンシング制御弁（以下、ＬＳ制御弁という）４１とを有している。ＬＳ制御弁には制御圧として油圧ポンプ１の吐出圧力と信号ライン３７の最高負荷圧とが対向して作用する。ポンプ吐出圧力が最高負荷圧力とバネ４１ａの設定値（目標ＬＳ差圧）との合計圧力よりも高くなると、アクチュエータ４０の油圧室４０ａを油圧ポンプ１の吐出油路１ｂに接続し、油圧室４０ａに高圧を導くことでピストン４０ｂをバネ４０ｃの力に打ち勝って図示左方に移動し、斜板１ａの傾転を減少させて油圧ポンプ１の吐出流量を減らす。逆に、ポンプ吐出圧力が最高負荷圧力とバネ４１ａの設定値（目標ＬＳ差圧）との合計圧力よりも低くなると、アクチュエータ４０の油圧室４０ａをタンク１９に接続し、油圧室４０ａを減圧することでバネ４０ｃの力でピストン４０ｂを図示右方に移動し、斜板１ａの傾転を増加させて油圧ポンプ１の吐出流量を増やす。このようなＬＳ制御弁の動作により、ポンプ吐出圧力が最高負荷圧力よりバネ４１ａの設定値（目標ＬＳ差圧）だけ高くなるように油圧ポンプ１の吐出流量が制御される。
【００４３】
圧力補償弁１２〜１６は、それぞれ、方向切換弁７〜１１の上流側の圧力を閉じ方向に作用させ、方向切換弁７〜１１の下流側の圧力である検出ライン２０〜２４の圧力（負荷圧）を開け方向に作用させると共に、信号ライン３７に導出した最高負荷圧力を閉じ方向に作用させ、油圧ポンプ１の吐出圧力を開け方向に作用させ、これにより上記のようにＬＳ制御された油圧ポンプ１の吐出圧力と最高負荷圧力との差圧（以下、適宜ＬＳ制御差圧という）を目標補償差圧としてそれぞれの方向切換弁７〜１１の前後差圧を制御するようになっている。
【００４４】
圧力補償弁１２〜１６に作用するそれぞれの方向切換弁７〜１１の上流側の圧力は信号ライン５０ａ〜５０ｅにより取り出され、方向切換弁７〜１１の下流側の圧力である検出ライン２０〜２４の圧力（負荷圧）は信号ライン５１ａ〜５１ｅにより取り出され、信号ライン３７の最高負荷圧力は信号ライン５２及び５２ａ〜５２ｅにより取り出され、油圧ポンプ１の吐出圧力は信号ライン５３及び５３ａ〜５３ｅにより取り出される。圧力補償弁１３〜１６において、信号ライン５２b〜５２ｅにより取り出された最高負荷圧力は油室１３ａ〜１６ａに負荷され、信号ライン５３ｂ〜５３ｅにより取り出された油圧ポンプ１の吐出圧力は油室１３ｂ〜１６ｂに負荷され、上記の目標補償差圧を設定する。圧力補償弁１２の目標補償差圧を設定する油室については後述する。
【００４５】
また、圧力補償弁１２は、方向切換弁７の上流側の圧力を閉じ方向に作用させ、方向切換弁７の下流側の圧力である検出ライン２０の圧力（旋回モータ２の負荷圧）を開け方向に作用させるときに、旋回モータ２の負荷圧が上昇すると、方向切換弁７を通過する圧油の流量を制限するよう目標補償差圧を小さくする負荷依存特性を有する構成になっていると共に、目標補償差圧の設定側である開け方向作用側に下限設定バネ５５を有している。この下限設定バネ５５は他のセクションの圧力補償弁１３〜１６の目標補償差圧がバネ５５の設定値よりも低くなったときにのみ圧力補償弁１２のスプールに作用し、目標補償差圧がその設定値以下に小さくならないよう下限を設定するものである。
【００４６】
圧力補償弁１２の構造を図２に示す。
【００４７】
図２において、圧力補償弁１２は、第１ボディ３０１ａと第２ボディ３０１ｂの２つのボディを有し、これらボディは適宜ボルト締め等の方法で（図示せず）一体に組付けられている。第１ボディ３０１ａには小径穴３２１と、この小径穴３２１に続く中径穴３２２とが設けられ、小径穴３２１に直径ｄ1の第１スプール３１１が摺動可能に嵌合し、中径穴３２２に直径ｄ3（＞ｄ1）の第２スプール３１２が摺動可能に嵌合している。第２ボディ３０１ｂには前記中径穴３２２に続く大径穴３２３と、この大径穴３２３に続く、前記小径穴３２１と同径の小径穴３２５とが設けられ、大径穴３２３及び小径穴３２５に第３スプール３１０が摺動可能に嵌合し、この第３スプール３１０は大径穴３２３に摺動可能に嵌合する直径ｄ2（＞ｄ3）の第１及び第２の大径部３１３，３１４と、小径穴３２５に摺動可能に嵌合する直径ｄ1の小径部３１５とを有している。
【００４８】
小径穴３２１の端面には凸部３２１ａが設けられ、凸部３２１ａの周囲に油室３３１が形成されると共に、第１スプール３１１の端面には凸部３２１ａを受け入れる凹部３１１ａが設けられ、凸部３２１ａの端面と凹部３１１ａの底部との間に上記各スプールを閉じ方向に押す初期位置保持用の弱いスプリング３５０を配している。また、スプリング３５０が配された室は凸部３２１ａ内に形成された通路３２１ｂを介して外部の油室３３１と連通している。
【００４９】
油室３３１の凸部３２１の周囲に上記の下限設定バネ５５が配され、第１スプール３１１の端面に向き合っている。この下限設定バネ５５は、図示の初期位置では第１スプール３１１の端面に向き合っているだけでそれから離れており、上記各スプールを開き方向に押す力は生じない。
【００５０】
また、ボディ３０１ａにはポンプポート３４１及び負荷圧ポート３４２が形成され、ボディ３０１ｂにはタンクポート３４３、出口ポート３４４、入口ポート３４５、最高負荷圧力ポート３４６が形成されている。ポンプポート３４１は、油圧ポンプ１の吐出圧力の信号ライン５３ａと連通しかつ油室３３１に開口し、負荷圧ポート３４２は、負荷圧の信号ライン５１ａに連通しかつ小径穴３２１と中径穴３２２の接続部に形成した油室３３２に開口している。また、タンクポート３４３は、タンク１９に連通しかつ第２スプール３１２と第３スプール３１０との当接部を囲む大径穴３２３に設けた油室３３３に開口し、出口ポート３４４は、ロードチェック弁１７ａに接続されかつ第１及び第２のスプール大径部３１３，３１４間の大径穴３２３に設けた油室３２８に開口し、入口ポート３４５は、ポンプ吐出油路１ｂと連通しかつ第３スプール３１０の第２の大径部３１４に設けた開閉可能な絞り部３１６の入側に開口し、最高負荷圧力ポート３４６は、最高負荷圧力の信号ライン５２ａと連通しかつ第３スプール３１０の第２の大径部３１４と小径部３１５との連続部が位置する大径穴３２３の部分に設けた油室３３６開口している。
【００５１】
また、小径部３１５と小径穴端面３３０間に、第３スプール３１０内に設けたパイロット油路５０ａを介して出口ポート３４４の油室３２８と連通する油室３３４を設けている。
【００５２】
第１ボディ３０１ａと第２ボディ３０１ｂは適宜ボルト締め等の方法で（図示せず）一体に組付けてボディ３０１を組成するが、この際第１ボディ３０１ａ側中径穴３２２と第２ボディ３０１ｂ側大径穴３２３とが芯ずれしていても、第２スプール３１２と第３スプール３１０は別部品で単に当接しているだけであることから、作動上の問題はない。
【００５３】
以上の構成により、圧力補償弁１２は閉じ方向に出口ポート３４４の出口圧力（Ｐz）をパイロット油路５０ａを介して油室３３４内の小径部３１５の端面３４０の受圧面積Ｂ1に、最高負荷圧力ポート３４６の最高負荷圧力（ＰLmax）を油室３３６内の第２の大径部３１４の断面積から小径部３１５の断面積を差し引いた段差部の受圧面積Ｂ2に、それぞれ作用させる。また、圧力補償弁１２は開く方向にポンプポート３４１を介してポンプ吐出圧力（Ｐs）を油室３３１内の第１スプール３１１の端面の受圧面積Ｂ1に、負荷圧力ポート３４２の負荷圧力（ＰL）を油室３３２内の第２スプール３１２の断面積から第１スプール３１１の断面積Ｂ1を差し引いた段差部の受圧面積Ｂ3に、それぞれ作用させる。なお、油室３３３内の第１の大径部３１３の断面積から第２スプール３１２の断面積を引いた段差部の受圧面積は、油室３３がタンクポート３４３によりタンク１９に通じているため、前記各スプールを開閉させる作用力は働らかない。
【００５４】
そして、上記受圧面積Ｂ2と第１スプール３１１の受圧面積Ｂ1とをほぼ同じとし（Ｂ1＝Ｂ2）、加えて受圧面積Ｂ3は第１スプールの受圧面積Ｂ1（＝Ｂ2）よリ小にし（Ｂ1＞Ｂ3）、旋回モータ２の負荷圧（ＰL）の増加に応じてその旋回モータ２に通じる方向切換弁７の通過流量を減少する負荷依存特性を持たせたものである。
【００５５】
即ち、第１スプール３１１、第２スプール３１２及び第３スプール３１３の油圧バランスを考えると、Ｂ1Ｐs−Ｂ2ＰLmaxに対しＢ1Ｐz−Ｂ3ＰLがつり合うことで圧力補償弁１２は機能するため、以下の式が成り立つ。
【００５６】
Ｂ1Ｐs−Ｂ2ＰLmax＝Ｂ1Ｐz−Ｂ3ＰL
Ｂ1＝Ｂ2より、
Ｂ1（Ｐs−ＰLmax）＝Ｂ2Ｐz−Ｂ3ＰL
Ｐs−ＰLmaxはＬＳ制御された油圧ポンプ１の吐出圧力Ｐsと最高負荷圧力ＰLmaxとの差圧（ＬＳ制御差圧）であるので、これをΔＰcとすると、
Ｂ1ΔＰc＝Ｂ2Ｐz−Ｂ3ＰL …（１）
方向切換弁７の前後差圧をΔＰとすると、
ΔＰ＝Ｐz−ＰL
となる。また、（１）式を変形して、
Ｂ1ΔＰc＋（Ｂ3−Ｂ2）ＰL＝Ｂ2（Ｐz−ＰL）
よって、

Figure 0003853123
ここで、Ｂ1／Ｂ2＝α、Ｂ3／Ｂ2＝βとおくと、
ΔＰ＝Ｐz−ＰL＝αΔＰc−（１−β）ＰL …（３）
即ち、Ｂ2＝Ｂ3であれば（Ｂ2とＢ3に面積差がなければ）、
ΔＰ＝αΔＰc
でΔＰはΔＰc（ＬＳ制御差圧）だけで決まるが、Ｂ2≠Ｂ3で面積差があるため、ΔＰはその面積差により負荷圧ＰLの影響を受け、負荷圧ＰLが増加するに従ってΔＰを小さくし方向切換弁７の通過流量を減少する負荷依存特性を有している。
【００５７】
図３に圧力補償弁１２の負荷依存特性を示す。図３の横軸は負荷圧であり、ＰLで表し、縦軸は目標補償差圧であり、ΔＰvで表している。点線は旋回モータ２のセクション（以下、旋回セクションという）以外の圧力補償弁１３〜１６の目標補償差圧を参考に示している。旋回セクション以外の圧力補償弁１３〜１６はそれらのアクチュエータ３〜６の負荷圧ＰLが増加しても、目標補償差圧ΔＰvはＬＳ制御差圧ΔＰcに保たれるが、旋回セクションの圧力補償弁１２は、負荷圧ＰLが増加すると負荷圧ＰLの増加に従って目標補償差圧ΔＰvが小さくなる。
【００５８】
図４に、圧力補償弁１２に負荷依存特性がないと仮定した場合の下限設定バネ５５による目標補償差圧の下限設定機能を示す。図４の横軸は、方向切換弁７とその他の方向切換弁８〜１１が要求する流量（バルブ要求流量）の総和であり、Ｑrで表している。これは方向切換弁７〜１１を切り換え操作するための図示しない操作レバー装置のレバー操作量の合計、即ち旋回モータ２及びそのアクチュエータの全要求流量に対応する。縦軸は圧力補償弁１２及びその他の圧力補償弁１３〜１６に設定される目標補償差圧ΔＰvである。また、下限設定バネ５５の設定差圧（目標補償差圧の下限値）をＰbとする。
【００５９】
旋回モータ２とその他のアクチュエータを同時に駆動する旋回複合動作時、方向切換弁７とその他の方向切換弁８〜１１のバルブ要求流量の総和Ｑrが油圧ポンプ１の最大吐出流量Ｑｐｍａｘよりも少なく、油圧ポンプ１の吐出流量がサチュレーション状態にないときは、圧力補償弁１２を含め全ての圧力補償弁の目標補償差圧ΔＰvはＬＳ制御差圧ΔＰcで一定である。
【００６０】
バルブ要求流量の総和Ｑrが油圧ポンプ１の最大吐出流量Ｑｐｍａｘを超え、油圧ポンプ１の吐出流量がサチュレーション状態になると、ＬＳ制御差圧ΔＰcが旋回セクションの圧力補償弁１２の下限設定バネ５５の設定差圧Ｐbに低下するまでは、全ての圧力補償弁の目標補償差圧ΔＰvはＬＳ制御差圧ΔＰcの低下と共に小さくなり、ＬＳ制御差圧ΔＰcが下限設定バネ５５の設定差圧Ｐbまで低下すると、それ以降は旋回セクションの圧力補償弁１２の目標補償差圧ΔＰvは下限設定バネ５５の設定差圧Ｐbに保持され、それ以下には小さくならず、旋回セクション以外の圧力補償弁の目標補償差圧ΔＰvは、ＬＳ制御差圧ΔＰcの低下と共に小さくなり続ける。
【００６１】
図中、太線の破線は旋回セクションを含む複合動作時の旋回セクション以外の圧力補償弁１３〜１６の目標補償差圧ΔＰvの変化であり、細線の破線は旋回セクションを含まない複合動作時の圧力補償弁１３〜１６の目標補償差圧ΔＰvの変化である。旋回セクションを含む複合動作時の旋回セクション以外の圧力補償弁１３〜１６の目標補償差圧ΔＰvは、旋回セクションの圧力補償弁１２の目標補償差圧ΔＰvが下限設定バネ５５の設定差圧Ｐbより小さくならないことから、旋回セクションを含まない複合動作時の圧力補償弁１３〜１６の目標補償差圧ΔＰvよりも低下の度合いが大きくなる。
【００６２】
以上の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る。図５に油圧ショベルの外観を示す。図５において、油圧ショベルは下部走行体２００、上部旋回体２０１、フロント作業機２０２を有し、上部旋回体２０１は下部走行体２００上に軸Ｏを中心に旋回可能であり、フロント作業機２０２は上部旋回体２０１の前部で上下動可能である。フロント作業機２０２はブーム２０３、アーム２０４、バケット２０５を有する多関節構造であり、ブーム２０３はブームシリンダ２０６により、アーム２０４はアームシリンダ２０７により、バケット２０５はバケットシリンダ２０８によりそれぞれ軸Ｏを含む平面内を回転駆動される。図１に示す旋回モータ２は上部旋回体２０２を下部走行体２００上に旋回駆動するアクチュエータであり、アクチュエータ３〜６のうちの３つがブームシリンダ２０６、アームシリンダ２０７、バケットシリンダ２０８として用いられる。
【００６３】
以上において、圧力補償弁１３〜１６の信号ライン５２ｂ〜５２ｅ，５３ｂ〜５３ｅにつながる油室１３ａ〜１６ａ，１３ｂ〜１６ｂは、複数の圧力補償弁１２〜１６のうち、旋回モータ２に係わる旋回セクション以外の圧力補償弁１３〜１６に設けられ、油圧ポンプ１の吐出圧力と複数のアクチュエータ２〜６の最高負荷圧との差圧を目標補償差圧として設定する第１手段を構成し、圧力補償弁１２の信号ライン５２ａ，５３ａにつながる油室３３６（受圧面積Ｂ2＝Ｂ1）及び油室３３１（受圧面積Ｂ1）は、旋回セクションの圧力補償弁１２に設けられ、その目標補償差圧を設定する第２手段を構成し、圧力補償弁１２の信号ライン５０ａ，５１ａにつながる油室３３４（受圧面積Ｂ1＞Ｂ3）及び油室３３２（受圧面積Ｂ3）は、複数の圧力補償弁１２〜１６のうち、少なくとも旋回セクションの圧力補償弁１２に設けられ、旋回モータ２の負荷圧が上昇すると、上記第２手段で設定された目標補償差圧を小さくし、旋回セクションの圧力補償弁１２に負荷依存特性を持たせる第３手段を構成し、圧力補償弁１２の下限設定バネ５５は、旋回セクションの圧力補償弁１２に設けられ、上記第２手段で設定され、上記第３手段で補正される目標補償差圧の下限を設定する第４手段を構成する。
【００６４】
また、本実施形態において、上記第２手段（油室３３１，３３６）は、第１手段（油室１３ａ〜１６ａ，１３ｂ〜１６ｂ）と同様、油圧ポンプ１の吐出圧力と複数のアクチュエータ２〜６の最高負荷圧との差圧を目標補償差圧として設定する手段であり、上記４手段（下限設定バネ５５）は、第２手段（油室３３１，３３６）で設定された目標補償差圧自体の低下と第３手段（油室３３２，３３４）で与えられた負荷依存特性による目標補償差圧の低下の両方に対して下限設定手段として機能する。
【００６５】
更に、上記第４手段（下限設定バネ５５）は、第２手段（油室３３１，３３６）で設定され、第３手段（油室３３２，３３４）で補正される目標補償差圧が所定値に達すると、旋回セクションの圧力補償弁１２のスプール３１１に開け方向の付勢力を付与する付勢手段である。
【００６６】
以上のように構成した本実施形態の動作を説明する。
１．旋回単独動作時
図６に、旋回用の方向切換弁７を操作し、旋回モータ２を単独で駆動する旋回単独動作時の旋回用の圧力補償弁１２の挙動をタイムチャートで示す。
【００６７】
旋回単独動作の起動時は、上部旋回体２０１の慣性負荷特有の負荷圧の上昇がある。この負荷圧の上昇は、旋回モータ２に設けられているオーバロードリリーフ弁６０ａ又は６０ｂなる安全弁により制限される。この状態では、旋回モータ２に供給された圧油は、安全弁６０ａ又は６０ｂよりタンクに放出される。
【００６８】
従来の一般的な圧力補償弁では、この安全弁からの圧油の放出により慣性負荷である上部旋回体２０１の加速感を調整していた。しかし、この場合は、起動時での旋回モータの消費流量が少ないことから、ほとんどの圧油がタンクに放出され、エネルギーロスとなる。また、油圧ポンプのＬＳ制御と圧力補償弁の流量補償機能とのバランスが取り難く、旋回操作性はギクシャクとしたものになる。
【００６９】
これに対し、本実施形態では、旋回セクションの圧力補償弁１２は上記のように負荷依存特性があるため、そのような問題は生じない。
【００７０】
まず、旋回用の操作レバー装置の操作レバーが操作されない起動前の状態では、圧力補償弁１２の目標補償差圧ΔＰvはＬＳ制御差圧ΔＰcに制御されている（ｔ0〜ｔ1）。
【００７１】
操作レバーを操作して旋回モータ２を起動すると、起動と同時に慣性負荷により負荷圧ＰLが上昇する（ｔ1）。
【００７２】
圧力補償弁１２の負荷依存特性により、目標補償差圧ΔＰvはＬＳ制御差圧ΔＰcから下がり、下限設定バネ５５の設定差圧Ｐbで下げ止まる（ｔ1）。旋回モータ２への供給流量Ｑaはバネ５５の設定差圧Ｐb相当の流量に制御される。下限設定バネ５５がない場合は、目標補償差圧ΔＰvはＰbより更に低い圧力まで下がる（０にはならない）。
【００７３】
上部旋回体２０１が回転を始め、旋回速度が上昇すると、旋回モータ２の消費流量と旋回モータ２への供給流量Ｑaがバランスし、負荷圧が徐々に低下する。その結果、圧力補償弁１２の目標補償差圧ΔＰvも上昇する（ｔ2）。
【００７４】
旋回モータ２の消費流量と供給流量Ｑaがバランスしない場合は、負荷圧ＰLの上昇又は低下となって旋回セクションの圧力補償弁１２にフィードバックされる。圧力補償弁１２の負荷圧依存特性により、供給流量Ｑaが多すぎた場合は負荷圧ＰLが高くなり、その結果、供給流量Ｑaは圧力補償弁１２により制限される。逆に、供給流量Ｑaが不足した場合は、負荷圧ＰLが低下し、供給流量Ｑaは圧力補償弁１２により増加される。この圧力補償弁１２の微調整により、旋回モータ２は従来のＬＳ制御で発生するようなハンチングを起こすことなく、緩やかに加速する。
【００７５】
本来の供給流量に達した時点で定常状態となり（ｔ3）、負荷圧ＰLは回転抵抗分の圧力となる。
２．旋回定常回転中の他のアクチュエータの起動
図７に、旋回単独で定常回転しているところに、他のアクチュエータ、例えばブームシリンダを起動し、複合動作した場合の各セクションの圧力補償弁の挙動をタイムチャートで示す。ブームシリンダはアクチュエータ３であるとする。
【００７６】
旋回単独定常回転時、旋回モータ２の負荷圧ＰLは定常回転に必要な圧力まで下がっており、圧力補償弁１２の目標補償差圧ΔＰvはほぼＬＳ制御差圧ΔＰcに制御されている（ｔ0〜ｔ1）。
【００７７】
ブーム用の操作レバー装置の操作レバーを追加操作した場合、旋回モータ２及びブームシリンダ３が合わせて要求する流量が、油圧ポンプ１が供給可能な最大吐出流量を超え、サチュレーションが発生すると、要求流量Ｑrに対する供給不足分に比例したＬＳ制御差圧ΔＰcの低下により各圧力補償弁１２，１３の目標補償差圧ΔＰvが下がり、流量の再分配が発生する（ｔ1）。
【００７８】
ここで、サチュレーションの度合いが大きい場合は、目標補償差圧ΔＰvは大きく低下するが、旋回セクションの圧力補償弁１２の目標補償差圧ΔＰvの低下は下限設定バネ５５の設定差圧Ｐbで制限される。このため、ブームセクションの圧力補償弁１３の目標補償差圧ΔＰvは、旋回側の目標補償差圧ΔＰvの低下が制限された分だけ更に低くなる。
【００７９】
結果として、旋回を含んだ複合動作時に、ある程度旋回モータ２へ優先的に圧油を供給することが可能となる。この機能により、サチュレーション状態時に旋回モータ２の他のアクチュエータに対する独立した操作性を実現でき、複合動作時の旋回の速度変化を抑え、旋回操作性を確保することが可能となる。
【００８０】
比較例として、旋回を含まない複合動作では、サチュレーションによるＬＳ制御差圧ΔＰcの低下により目標補償差圧ΔＰvは同じ値に低下し、供給流量Ｑaも同じ値に低下する（複合動作に係わる方向切換弁の開口面積は同一と仮定）。旋回セクションの圧力補償弁１２に下限設定バネ５５がない場合（特開平１０−８９３０４号の場合）の旋回を含む複合動作でも同様であり、下限設定バネ５５を設けることにより、その場合と比較してもΔΔＰv1，ΔＱa1だけ旋回セクションの目標補償差圧ΔＰv及び供給流量Ｑaの低下が抑えられ、旋回モータ２へ優先的に圧油が供給され、複合動作時の旋回の速度変化を抑えられる。
【００８１】
図８は上記複合動作における油圧ポンプ１の吐出流量のサチュレーションの度合いが小さい場合である。
【００８２】
サチュレーションの度合いが小さい場合は、目標補償差圧ΔＰvの低下は下限設定バネ５５の設定差圧Ｐb以上にとどまる。この場合、旋回・ブームとも同一の目標補償差圧ΔＰv及び流量Ｑaに低下する（旋回及びブームセクションの方向切換弁７，８の開口面積は同一と仮定）。
【００８３】
このように下限設定バネ５５の設定により、旋回の優先の度合いをサチュレーションの度合いにより設定することが可能になる。
３．旋回と他のアクチュエータとの同時起動
図９に、旋回起動時に同時に他のアクチュエータ、例えばブームシリンダを起動した複合動作時の各セクションの圧力補償弁の挙動をタイムチャートで示す。この場合もブームシリンダはアクチュエータ３であるとする。
【００８４】
まず、旋回用及びブーム用の操作レバー装置の操作レバーが操作されない起動前の状態では、圧力補償弁１２，１３の目標補償差圧ΔＰvはＬＳ制御差圧ΔＰcに制御されている（ｔ0〜ｔ1）。
【００８５】
旋回及びブーム用の操作レバーを同時操作して旋回モータ２及びブームシリンダ３を同時起動したとき、旋回とブームを合わせた要求流量が油圧ポンプ１の最大吐出流量を超え、サチュレーションが発生すると、要求流量Ｑrに対する供給不足分に比例したＬＳ制御差圧ΔＰcの低下により各圧力補償弁１２〜１６の目標補償差圧ΔＰvが下がり、流量の再分配が発生する（ｔ1）。
【００８６】
この場合も、旋回セクションの圧力補償弁１２の負荷依存特性による微調整により、旋回モータ２は従来のＬＳ制御で発生するようなハンチングを起こすことなく、緩やかに加速する。
【００８７】
また、サチュレーションの度合いが大きい場合は、目標補償差圧ΔＰvは大きく低下する。更に、旋回セクションの圧力補償弁１２については、旋回モータ２の起動と同時に慣性負荷により旋回モータ２の負荷圧ＰLが上昇するため、圧力補償弁１２の負荷依存特性によっても目標補償差圧ΔＰvの低下がある。この圧力補償弁１２の目標補償差圧ΔＰvの低下は下限設定バネ５５の設定差圧Ｐbによって制限される。このため、ブームセクションの圧力補償弁１３の目標補償差圧ΔＰvは、旋回側の目標補償差圧ΔＰvの低下が制限された分だけ更に低くなる。
【００８８】
結果として、油圧ポンプ１の吐出流量はある程度旋回モータ２へ優先的に供給され、ブームシリンダ３に比べ、旋回速度が極端に遅くなることなく、旋回操作性を維持することができる。る。
【００８９】
比較例として、旋回を含まない複合動作では、図９に破線で示すように、サチュレーションによるＬＳ制御差圧ΔＰcの低下により目標補償差圧ΔＰvは同じ値に低下し、供給流量Ｑaも同じ値に低下する（複合動作に係わる方向切換弁の開口面積は同一と仮定）。
【００９０】
旋回セクションの圧力補償弁１２に下限設定バネ５５がない場合（特開平１０−８９３０４号の場合）の旋回を含む複合動作では、サチュレーションによるＬＳ制御差圧ΔＰcの低下と圧力補償弁１２の負荷依存特性とによって目標補償差圧ΔＰvは、図９に二点鎖線で示すように極端に低下し、供給流量Ｑaも極端に減少する。本実施形態では、この圧力補償弁１２の目標補償差圧ΔＰvの低下は下限設定バネ５５の設定差圧Ｐbによって制限される。このため、バネ５５を設けない場合と比較してΔΔＰv2，ΔＱa2だけ旋回セクションの目標補償差圧ΔＰv及び供給流量Ｑaの低下が抑えられる。この機能により、複合動作時に、他のアクチュエータに比べ、旋回速度が極端に遅くなることなく、旋回操作性を維持することができる。
【００９１】
図１０は上記複合動作における油圧ポンプ１の吐出流量のサチュレーションの度合いが小さい場合である。
サチュレーションの度合いが小さい場合、ブームセクションの圧力補償弁１３の目標補償差圧ΔＰvの低下は旋回セクションの圧力補償弁１２の下限設定バネ５５の設定差圧Ｐb以上にとどまる。旋回セクションの圧力補償弁１２の負荷依存性により、旋回セクションの目標補償差圧ΔＰvは下限設定バネ５５の設定差圧Ｐbまで低下する。
【００９２】
旋回速度が上昇するにつれ、旋回モータ２の負荷圧が低下し、旋回セクションの圧力補償弁１２の目標補償差圧ΔＰvが上昇する。最終的には旋回、ブームセクションとも同一の目標補償差圧ΔＰv及び供給流量Ｑaになる（旋回、ブームセクションの方向切換弁の開口面積は同一と仮定）（ｔ４）。
【００９３】
旋回セクションの圧力補償弁１２に下限設定バネ５５がない場合（特開平１０−８９３０４号の場合）は、図１０に二点鎖線で示すように、旋回セクションの圧力補償弁１２の目標補償差圧ΔＰvは、Ｐbより更に低い圧力まで下がり、旋回モータ２への供給流量Ｑaも、起動直後は大幅に低下する。下限設定バネ５５を設けることにより、その場合と比較して起動直後はΔΔＰv3，ΔＱa3だけ旋回セクションの目標補償差圧ΔＰv及び供給流量Ｑaの低下が抑えられる。したがって、この場合も、他のアクチュエータに比べ、旋回速度が極端に遅くなることなく、旋回操作性を維持することができる。
【００９４】
以上のように本実施形態によれば、旋回セクションの圧力補償弁１２の負荷依存特性により、旋回単独、複合のいずれの起動時にも、旋回操作性のギクシャク感がなく加速して定常状態に移行できる。また、旋回セクションの圧力補償弁１２に下限設定バネ５５を設け、油圧ポンプ１の吐出流量のサチュレーション時に旋回モータ２に優先的に圧油を供給するようにしたので、旋回単独動作から旋回複合動作への移行時旋回速度変化が抑えられ、逆の旋回複合から旋回単独動作への移行時にも同様であり、更に旋回複合の起動時に、他のアクチュエータに比べ旋回速度が極端に遅くならずに加速でき、優れた旋回操作性と旋回独立性を確保できる。また、別回路を設けることなく上記の機能を達成するので、コスト・スペースの増加や回路構成の複雑化の問題も生じない。
【００９５】
本発明の第２の実施形態を図１１〜図１４により説明する。図中、図１及び図２に示した部材と同等の部材荷は同じ符号を付している。本実施形態は、旋回優先バネを常時圧力補償弁のスプールに作用させるようにしたものである。
【００９６】
図１１において、旋回セクション以外の圧力補償弁１３〜１６は第１の実施形態のものと同じである。
【００９７】
旋回セクションの圧力補償弁１２Ａは、方向切換弁７Ａの上流側の圧力を閉じ方向に作用させ、方向切換弁７Ａの下流側の圧力である検出ライン２０〜２４の圧力（負荷圧）を開け方向に作用させると共に、信号ライン３７に導出した最高負荷圧力を閉じ方向に作用させ、油圧ポンプ１の吐出圧力を開け方向に作用させ、これによりＬＳ制御差圧（ＬＳ制御された油圧ポンプ１の吐出圧力と最高負荷圧力との差圧）を目標補償差圧として方向切換弁７Ａの前後差圧を制御するようになっていると共に、旋回モータ２の負荷圧が上昇すると、方向切換弁７Ａを通過する圧油の流量を制限するよう目標補償差圧を小さくする負荷依存特性を有する構成になっており、この点も第１の実施形態の圧力補償弁１２と同じである。
【００９８】
そして圧力補償弁１２Ａは、目標補償差圧の設定側である開け方向作用側に旋回優先バネ５５Ａを有し、この旋回優先バネ５５Ａは、圧力補償弁１２Ａの動作中、常時圧力補償弁１２Ａのスプールに作用し、上記のＬＳ制御差圧による目標補償差圧に加算される旋回優先用の一定の補助的な目標補償差圧を設定している。即ち、圧力補償弁１２Ａの目標補償差圧は、旋回セクション以外の圧力補償弁１３〜１６よりも旋回優先バネ５５Ａによる設定分だけ大きくなっている。
【００９９】
また、旋回セクションの方向切換弁７Ａは、その圧力補償弁１２Ａの大きめの目標補償差圧の設定に対応して、油圧ポンプ１の吐出流量がサチュレーション状態にないときに設計通りの流量特性が得られるよう、メータインの可変絞り５７ａ，５７ｂの開口面積を通常より小さく設定している。
【０１００】
図１２にその関係を示す。図中、Ｍ1は方向切換弁７Ａのスプールストロークに対するメータインの可変絞り５７ａ，５７ｂの開口面積の変化（開口面積特性）であり、Ｍ2は圧力補償弁に旋回優先バネ５５Ａを用いない、定格条件での方向切換弁（例えば図１に示す第１の実施形態における方向切換弁７）のスプールストロークに対するメータイン可変絞りの開口面積の変化（開口面積特性）である。Ｍ2よりもＭ1の方が同じスプールストロークに対して開口面積が小さくなるように設定されている。
【０１０１】
圧力補償弁１２Ａの構造を図１３に示す。図１３において、第１ボディ３０１ａには端面３２０を有する小径穴３２１が形成されており、この小径穴３２１の端面３２０の部分の油室３３１Ａにおいて、小径穴３２１に嵌合する第１スプール３１１と小径穴３２１の端面３２０との間に第１スプール３１１、第２スプール３１２、第３スプール３１０を閉じ方向に押す上記の旋回優先バネ５５Ａが配されている。油室３３１Ａ，３３２，３３４，３３６内の受圧面積Ｂ1，Ｂ3，Ｂ1，Ｂ2の関係は第１の実施形態の図２に示す油室３３１，３３２，３３４，３３６内の受圧面積Ｂ1，Ｂ3，Ｂ1，Ｂ2の関係と同じである。また、圧力補償弁１２Ａのその他の構成も図２に示す第１の実施形態のものと同じである。
【０１０２】
圧力補償弁１２Ａにおける旋回優先バネ５５Ａの動作原理を説明する。
【０１０３】
第１の実施形態の圧力補償弁１２における下限設定バネ５５は、目標補償差圧が既定値以下に小さくならないよう目標補償差圧に下限を設定していた。この目標補償差圧の下限の値を前述のＰbとすると、本実施形態では、旋回優先バネ５５Ａを常時スプールに作用させ、その下限値Ｐb相当の目標補償差圧をＬＳ制御差圧による目標補償差圧に加算されるものとして設定する。その結果、圧力補償弁１２Ａの目標補償差圧は他の圧力補償弁１３〜１６よりＰb分だけ大きくなる。即ち、
圧力補償弁１３〜１６の目標補償差圧：Ｐs−ＰLmax
圧力補償弁１２Ａの目標補償差圧 ：Ｐs−ＰLmax＋Ｐb
このように圧力補償弁１２Ａの目標補償差圧を設定すると、旋回セクションの方向切換弁のメータイン可変絞りの開口面積を今までと同じ大きさに設定したのでは、旋回モータ２にだけＰb分の流量が多く流れることになる。従って、旋回モータ２に今までと同じ流量が流れるように、旋回セクションの方向切換弁のメータイン可変絞りの開口面積をＰb分小さくする必要がある。
【０１０４】
即ち、本来の定格条件の目標補償差圧での旋回の方向切換弁の開口面積をＡsとし、方向切換弁７Ａのメータイン可変絞りの開口面積をＡsoとすると、
Ａso＝Ａs√（（Ｐs−ＰLmax）／（Ｐs−ＰLmax＋Ｐb））
となる。
【０１０５】
このような圧力補償弁１２Ａ及び方向切換弁７Ａを用いた場合のサチュレーション時の旋回モータ２への供給流量の変化を他のアクチュエータと比較する。他のアクチュエータに係わる方向切換弁の開口面積を定格条件の目標補償差圧での旋回の方向切換弁の開口面積と同じＡsとし、旋回モータ２への供給流量をＱaとし、他のアクチュエータへの供給流量をＱbとすると、Ｑa，Ｑbはそれぞれ次のように表せる。
【０１０６】
Figure 0003853123
ここで、Ａs√（（Ｐs−ＰLmax）／（Ｐs−ＰLmax＋Ｐb））は定格条件での値（常数）である。
【０１０７】
定格条件を下記のように設定する。
【０１０８】
Figure 0003853123
これらの値を上記ＱbとＱaの式に代入する。
【０１０９】
Ｑb＝２１．９４√ΔＰc
Ｑa＝２１．９４×０．９１√（ΔＰc＋Ｐb）
上記のＱa，ＱbとＬＳ制御差圧ΔＰcとの関係を比較して示すと図１４のようになる。この図から分かるように、ＬＳ制御差圧ΔＰcが１５ｋｇｆ／ｃｍ２以下になると、即ち油圧ポンプ１の吐出流量が要求流量に満たないサチュレーション状態になると、旋回モータ２の供給流量Ｑaが旋回以外のアクチュエータの供給流量Ｑbよりも多くなり、旋回モータ２に優先的に圧油が供給される。また、その優先の度合い（流量の差）はＬＳ制御差圧ΔＰcが小さくなるに従って大きくなる。
【０１１０】
以上において、圧力補償弁１３〜１６の信号ライン５２ｂ〜５２ｅ，５３ｂ〜５３ｅにつながる油室１３ａ〜１６ａ，１３ｂ〜１６ｂは、複数の圧力補償弁１２〜１６のうち、旋回モータ２に係わる旋回セクション以外の圧力補償弁１３〜１６に設けられ、油圧ポンプ１の吐出圧力と複数のアクチュエータ２〜６の最高負荷圧との差圧を目標補償差圧として設定する第１手段を構成し、圧力補償弁１２Ａの信号ライン５２ａ，５３ａにつながる油室３３６（受圧面積Ｂ2＝Ｂ1）及び油室３３１Ａ（受圧面積Ｂ1）は、旋回セクションの圧力補償弁１２に設けられ、その目標補償差圧を設定する第２手段を構成し、圧力補償弁１２Ａの信号ライン５０ａ，５１ａにつながる油室３３４（受圧面積Ｂ1＞Ｂ3）及び油室３３２（受圧面積Ｂ3）は、複数の圧力補償弁１２〜１６のうち、少なくとも旋回セクションの圧力補償弁１２Ａに設けられ、旋回モータ２の負荷圧が上昇すると、上記第２手段で設定された目標補償差圧を小さくし、旋回セクションの圧力補償弁１２Ａに負荷依存特性を持たせる第３手段を構成し、圧力補償弁１２Ａの旋回優先バネ５５Ａは、旋回セクションの圧力補償弁１２Ａに設けられ、上記第２手段で設定され、上記第３手段で補正される目標補償差圧の下限を設定する第４手段を構成する。
【０１１１】
また、本実施形態において、上記第２手段（油室３３１Ａ，３３６）は、第１手段（油室１３ａ〜１６ａ，１３ｂ〜１６ｂ）と同様、油圧ポンプ１の吐出圧力と複数のアクチュエータ２〜６の最高負荷圧との差圧を目標補償差圧として設定する手段であり、上記４手段（旋回優先バネ５５）は、第２手段（油室３３１Ａ，３３６）で設定された目標補償差圧自体の低下と第３手段（油室３３２，３３４）で与えられた負荷依存特性による目標補償差圧の低下の両方に対して下限設定手段として機能する。
【０１１２】
更に、上記第４手段（旋回優先バネ５５）は、第２手段（油室３３１Ａ，３３６）で設定され、第３手段（油室３３２，３３４）で補正される目標補償差圧に常時補助的な値を付加する付勢手段であり、旋回セクションの方向切換弁７Ａは、そのメータイン可変絞り５７ａ，５７ｂの開口面積が、当該付勢手段で付加される補助的な値の目標補償圧相当分だけ、旋回セクション以外の方向切換弁８〜１１の開口面積より小さくなるように構成されている。
【０１１３】
したがって、本実施形態においても、旋回セクションの圧力補償弁１２Ａの負荷依存特性により、旋回単独、複合のいずれの起動時にも、旋回操作性のギクシャク感がなく加速して定常状態に移行できる。また、旋回セクションの圧力補償弁１２Ａに旋回優先バネ５５Ａを設け、油圧ポンプ１の吐出流量のサチュレーション時に旋回モータ２に優先的に圧油を供給するようにしたので、旋回単独動作から旋回複合動作への移行時旋回速度変化が抑えられ、逆の旋回複合から旋回単独動作への移行時にも同様であり、更に旋回複合の起動時に、他のアクチュエータに比べ旋回速度が極端に遅くならずに加速でき、優れた旋回操作性と旋回独立性を確保できる。また、別回路を設けることなく上記の機能を達成するので、コスト・スペースの増加や回路構成の複雑化の問題も生じない。
【０１１４】
本発明の第３の実施形態を図１５及び図１６により説明する。図中、図１及び図２に示した部材と同等の部材荷は同じ符号を付している。本実施形態は、旋回セクションの圧力補償弁にＬＳ制御差圧による目標補償差圧の設定を行わずに旋回優先性を与えたものである。
【０１１５】
図１５において、旋回セクション以外の圧力補償弁１３〜１６は第１の実施形態のものと同じである。
【０１１６】
また、旋回セクションの圧力補償弁１２Ｂは、方向切換弁７の上流側の圧力を閉じ方向に作用させ、方向切換弁７の下流側の圧力である検出ライン２０の圧力（旋回モータ２の負荷圧）を開け方向に作用させるときに、旋回モータ２の負荷圧が上昇すると、圧力補償弁１２Ｂを通過する圧油の流量を制限するよう目標補償差圧を小さくする負荷依存特性を有する構成になっており、この点は第１の実施形態の圧力補償弁１２と同じである。
【０１１７】
そして圧力補償弁１２Ｂは、目標補償差圧の設定側である開け方向作用側に通常の目標補償差圧を設定する手段、例えば設定バネ６０を有し、この設定バネ６０は、油圧ポンプ１の吐出流量がサチュレーション状態にないときのＬＳ制御差圧による目標補償差圧と同じ大きさの目標補償差圧を設定する構成となっている。即ち。ＬＳ制御差圧による目標補償差圧を設定する旋回セクション以外の圧力補償弁１３〜１６は、油圧ポンプ１の吐出流量がサチュレーション状態になると、サチュレーションの度合いに応じて目標補償差圧が小さくなるのに対して、旋回セクションの圧力補償弁１２Ｂは、サチュレーション状態になっても設定バネ６０により設定される目標補償差圧は実質的に不変であり、この目標補償差圧が負荷依存特性により変化する。
【０１１８】
また、圧力補償弁１２Ｂには、第１の実施形態と同様、圧力補償弁１２Ｂの目標補償差圧の下限を設定する下限設定バネ５５が設けられている。
【０１１９】
圧力補償弁１２Ｂの構造を図１６に示す。図１６において、図２に示した第１の実施形態における油室３３１，３３６はそれぞれ油室３３１Ｂ，３３６Ｂに置き換えられ、これら油室３３１Ｂ，３３６Ｂはそれぞれタンクポート３４１Ｂ，３４６Ｂを介してタンクに連通し、第１スプール３１１により与えられる油室３３１Ｂの受圧面積Ｂ1及び第３スプール３１０の第２の大径部３１４と小径部３２５間の段差部により与えられる油室３３６Ｂの受圧面積Ｂ2がそれぞれ第１スプール３１１及び第３スプール３１０に油圧力を作用しないように構成されている。また、第１スプール３１１の端面に形成された凹部３１１ａ内には初期位置保持用の弱いスプリング３５０に代え、上述した目標補償差圧を設定するバネ６０が配置されている。油室３３２，３３４に位置する受圧面積Ｂ3，Ｂ1の関係は第１の実施形態と同じであり（Ｂ1＞Ｂ3）、これにより旋回モータ２の負荷圧（ＰL）の増加に応じて旋回モータ２に通じる方向切換弁７の通過流量を減少する負荷依存特性を持たせている。
【０１２０】
以上において、圧力補償弁１３〜１６の信号ライン５２ｂ〜５２ｅ，５３ｂ〜５３ｅにつながる油室１３ａ〜１６ａ，１３ｂ〜１６ｂは、複数の圧力補償弁１２〜１６のうち、旋回モータ２に係わる旋回セクション以外の圧力補償弁１３〜１６に設けられ、油圧ポンプ１の吐出圧力と複数のアクチュエータ２〜６の最高負荷圧との差圧を目標補償差圧として設定する第１手段を構成し、圧力補償弁１２Ｂの設定バネ６０は、旋回セクションの圧力補償弁１２Ｂに設けられ、その目標補償差圧を設定する第２手段を構成し、圧力補償弁１２Ｂの信号ライン５０ａ，５１ａにつながる油室３３４（受圧面積Ｂ1＞Ｂ3）及び油室３３２（受圧面積Ｂ3）は、複数の圧力補償弁１２〜１６のうち、少なくとも旋回セクションの圧力補償弁１２Ｂに設けられ、旋回モータ２の負荷圧が上昇すると、上記第２手段で設定された目標補償差圧を小さくし、旋回セクションの圧力補償弁１２Ｂに負荷依存特性を持たせる第３手段を構成し、圧力補償弁１２の下限設定バネ５５は、旋回セクションの圧力補償弁１２に設けられ、上記第２手段で設定され、上記第３手段で補正される目標補償差圧の下限を設定する第４手段を構成する。
【０１２１】
また、本実施形態において、上記第２手段（設定バネ６０）は、油圧ポンプ１１の吐出圧力と複数のアクチュエータ２〜６の最高負荷圧との差圧により変化しない値を目標補償差圧として設定する手段であり、上記第４手段（下限設定バネ５５）は、第３手段（油室３３２，３３４）で与えれた負荷依存特性による目標補償差圧の低下に対して下限設定手段として機能する。
【０１２２】
更に、上記第４手段（下限設定バネ５５）は、第２手段（設定バネ６０）で設定され、第３手段（油室３３２，３３４）で補正される目標補償差圧が所定値に達すると、旋回セクションの圧力補償弁１２Ｂのスプール３１１に開け方向の付勢力を付与する付勢手段である。
【０１２３】
以上のように構成した本実施形態においては、設定バネ６０は、油圧ポンプ１の吐出流量がサチュレーション状態にないときのＬＳ制御差圧による目標補償差圧と同じ大きさの目標補償差圧を設定する構成となっているため、油圧ポンプ１の吐出流量がサチュレーションする前は、第１の実施形態と同様に複数のアクチュエータのそれぞれの要求流量の比で油圧ポンプ１の吐出流量を分配するよう目標補償差圧が設定され、かつ旋回セクションの圧力補償弁１２Ｂの負荷依存特性によりその目標補償差圧が補正される一方、油圧ポンプ１の吐出流量がサチュレーション状態になると、旋回セクション以外の圧力補償弁１３〜１６の目標補償差圧はＬＳ制御差圧の低下に応じて目標補償差圧が低下するのに対して、旋回セクションの圧力補償弁１２Ｂの設定バネ６０による目標補償差圧はサチュレーションの度合いによっては変化せず、圧力補償弁１２Ｂの目標補償差圧は負荷依存特性によってのみ変化しかつこの負荷依存特性による目標補償差圧の低下に対しては下限設定バネ５５が機能し、この場合も第１及び第２の実施形態と同様に旋回モータ２に優先的に圧油が供給されることとなる。
【０１２４】
したがって、本実施形態によっても、旋回セクションの圧力補償弁１２Ｂの負荷依存特性により、旋回単独、複合のいずれの起動時にも、旋回操作性のギクシャク感がなく加速して定常状態に移行できる。また、旋回セクションの圧力補償弁１２Ｂに下限設定バネ５５と設定バネ６０を設け、油圧ポンプ１の吐出流量のサチュレーション時及び負荷依存特性による目標補償差圧の低下時に旋回モータ２に優先的に圧油を供給するようにしたので、旋回単独動作から旋回複合動作への移行時旋回速度変化が抑えられ、逆の旋回複合から旋回単独動作への移行時にも同様であり、更に旋回複合の起動時に、他のアクチュエータに比べ、旋回速度が極端に遅くならずに加速でき、優れた旋回操作性と旋回独立性を確保できる。また、別回路を設けることなく上記の機能を達成するので、コスト・スペースの増加や回路構成の複雑化の問題も生じない。
【０１２５】
なお、上記実施形態では、方向切換弁の上流側に位置するビフォアオリフィスタイプの圧力補償弁を用いた例を示したが、方向切換弁の下流側に位置するアフタオリフィスタイプの圧力補償弁を用いても同等の効果を持つシステムを構成することが可能である。
【０１２６】
また、上記実施形態では、旋回セクションの圧力補償弁に優先性を持たせるよう目標補償差圧を制御する手段として下限設定バネ５５、旋回優先バネ５５Ａ、設定バネ６０を設けたが、方向切換弁の上下流の圧力が導かれる油室と同様に制御圧を導く油室を設け、油圧的な制御力を付与するようにしても良い。この場合、目的に応じて制御圧を変化させることにより、更に複雑で利点のある制御を行うことが可能となる。
【０１２７】
更に、上記実施形態では、油圧ポンプの吐出圧力と複数のアクチュエータの最高負荷圧との差圧を目標補償差圧として設定するのに、ポンプ吐出圧力と最高負荷圧とを圧力補償弁のスプールの対向端部に別々に導いたが、油圧ポンプの吐出圧力と複数のアクチュエータの最高負荷圧との差圧に対応した二次圧を発生する差圧発生弁を設け、その出力圧を圧力補償弁のスプールの開き方向の端部に導いても良い。
【０１２８】
【発明の効果】
本発明によれば、旋回制御系を含む油圧駆動装置において、旋回単独、複合のいずれの起動時にも、旋回操作性のギクシャク感がなく加速して定常状態に移行でき、しかも旋回単独動作から旋回複合動作への移行時又はその逆の場合の旋回速度変化が抑えられ、かつ複合の起動時に他のアクチュエータに比べ旋回速度が極端に遅くならずに加速でき、優れた旋回操作性と旋回独立性を確保できると共に、別回路を設けることによるコスト・スペースの増加や回路構成の複雑化の問題を生じないシステムとすることができる。
【図面の簡単な説明】
【図１】本発明の第１の実施形態による油圧駆動装置を示す回路図である。
【図２】旋回セクションの圧力補償弁の構造の詳細を示す断面図である。
【図３】旋回セクションの圧力補償弁の負荷依存特性を示す図である。
【図４】旋回セクションの圧力補償弁における旋回優先バネによる目標補償差圧の下限設定機能を示す図である。
【図５】本発明の油圧駆動装置が用いられる油圧ショベルの外観を示す図である。
【図６】旋回単独動作時における旋回セクションの圧力補償弁の目標補償差圧の変化を示すタイムチャートである。
【図７】旋回定常回転中に他のアクチュエータを起動した場合のサチュレーションの度合いが大きい場合の旋回セクションの圧力補償弁の動作を説明するタイムチャートである。
【図８】旋回定常回転中に他のアクチュエータを起動した場合のサチュレーションの度合いが小さい場合の旋回セクションの圧力補償弁の動作を説明するタイムチャートである。
【図９】旋回と他のアクチュエータの同時起動した場合のサチュレーションの度合いが大きい場合の旋回セクションの圧力補償弁の動作を説明するタイムチャートである。
【図１０】旋回と他のアクチュエータの同時起動した場合のサチュレーションの度合いが小さい場合の旋回セクションの圧力補償弁の動作を説明するタイムチャートである。
【図１１】本発明の第２の実施形態による油圧駆動装置を示す回路図である。
【図１２】旋回セクションの方向切換弁の開口面積特性を示す図である。
【図１３】旋回セクションの圧力補償弁の構造の詳細を示す断面図である。
【図１４】サチュレーション状態での旋回セクションの流量の優先特性を示す図である。
【図１５】本発明の第３の実施形態による油圧駆動装置を示す回路図である。
【図１６】旋回セクションの圧力補償弁の構造の詳細を示す断面図である。
【符号の説明】
１ 油圧ポンプ
２〜６ アクチュエータ（２：旋回モータ）
７〜１１ 方向切換弁
１２〜１６ 圧力補償弁
１３ａ〜１６ａ 油室（第１手段）
１３ｂ〜１６ｂ 油室（第１手段）
１８ ポンプ制御装置
２０〜２４ 検出ライン
２５〜２９ 信号ライン
３４〜３６ 信号ライン
３７ 信号ライン
４０ 傾転制御アクチュエータ
４１ ロードセンシング制御弁
５０ａ〜５０ｅ 信号ライン
５１ａ〜５１ｅ 信号ライン
５２，５２ａ〜５２ｅ 信号ライン
５３，５３ａ〜５２ｅ 信号ライン
５５ 下限設定バネ（第４手段；付勢手段）
６０ 設定バネ（第２手段）
２００ 下部走行体
２０１ 旋回体
２０２ フロント作業機
３３１ 油室（受圧面積Ｂ1）（第２手段）
３３２ 油室（受圧面積Ｂ3）（第３手段）
３３４ 油室（受圧面積Ｂ1＞Ｂ3）（第３手段）
３３６ 油室（受圧面積Ｂ2＝Ｂ1）（第２手段）
７Ａ 方向切換弁
１２Ａ 圧力補償弁
５５Ａ 旋回優先バネ（第４手段；付勢手段）
５７ａ，７５ｂ メータインの可変絞り
３３１Ａ 油室（受圧面積Ｂ1）（第２手段）
１２Ｂ 圧力補償弁[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a hydraulic drive device for a construction machine including a swing control system such as a hydraulic excavator, and in particular, when supplying pressure oil from a hydraulic pump to each of a plurality of actuators including a swing motor via respective direction switching valves. The present invention relates to a hydraulic drive apparatus that controls a discharge flow rate of a hydraulic pump by a load sensing system and controls a differential pressure across a direction switching valve by a respective pressure compensation valve.
[0002]
[Prior art]
Japanese Patent Application Laid-Open No. 60-11706 discloses a hydraulic drive device that controls a discharge flow rate of a hydraulic pump by a load sensing system (hereinafter referred to as LS system as appropriate). Japanese Patent Application Laid-Open No. 10-37907 discloses a hydraulic drive device for a construction machine including a turning control system that includes an LS system and realizes independence and operability of the turning control system. Further, as an open center type hydraulic drive device for a construction machine including a turning control system, an independence of the turning control system is realized as a three-pump system mounted on an actual machine. Further, Japanese Patent Application Laid-Open No. 10-89304 discloses a hydraulic drive device that controls the discharge flow rate of a hydraulic pump by using an LS system and has a pressure dependent valve having load dependent characteristics.
In the hydraulic drive device described in Japanese Patent Application Laid-Open No. 60-11706, the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators is set as the target compensation differential pressure for each of the plurality of pressure compensation valves. If the discharge flow rate of the hydraulic pump becomes less than the flow rate required by the plurality of directional control valves during a combined operation in which a plurality of actuators are driven simultaneously, this saturation state causes the hydraulic pump to By reducing the differential pressure between the discharge pressure and the maximum load pressure, each target compensation differential pressure of the pressure compensation valve becomes small, and the discharge flow rate of the hydraulic pump can be redistributed to the ratio of the flow rate required by each actuator.
[0003]
Japanese Patent Application Laid-Open No. 10-37907 discloses a three-pump system equipped with a hydraulic drive device and an actual machine, both of which are related to a swing section including a swing motor by an open center type independent circuit using an independent hydraulic pump. A separate circuit is constructed with the actuator to ensure the independence and operability of the turning control system.
[0004]
In the hydraulic drive device described in Japanese Patent Laid-Open No. 10-89304, for each of a plurality of pressure compensation valves, a hydraulic chamber of a closing direction action in which an inlet side pressure of a direction switching valve is guided among the hydraulic chambers of the pressure compensation valve. By making the pressure receiving area larger than the pressure receiving area of the hydraulic chamber of the opening direction action where the outlet side pressure of the direction switching valve is guided, the target compensation differential pressure of the pressure compensation valve is reduced with respect to the increase of the load pressure of each actuator. (Thresholding the pressure compensation valve) and having a load-dependent characteristic that reduces the supply flow rate to the actuator. This makes it easy to operate on both the low and high load sides, and does not cause hunting and is stable. Be able to work.
[0005]
[Problems to be solved by the invention]
However, the conventional hydraulic drive device has the following problems with respect to the turning control system.
[0006]
JP 60-11706 A: The following problems (1) (2)
JP-A-10-89304: The following problems (2) (3)
JP-A-10-37907: The following problems (4)
Open center type 3-pump system with actual equipment: The following problems (4)
▲ 1 ▼ Feeling of operability when turning alone
(2) Change in turning speed at the time of transition from turning single action to turning combined action or vice versa
(3) Extreme decrease in turning speed when turning combined
(4) Increase in cost / space and complexity of circuit configuration by providing separate circuits
[0007]
(1) Japanese Patent Laid-Open No. 60-11706
In the hydraulic drive apparatus provided with the LS system described in Japanese Patent Application Laid-Open No. 60-11706, when this is used for a swing control system, the swing control system is accompanied by an inertia load. It is difficult to balance the flow compensation function of the pressure compensation valve and the LS control as appropriate. This is because it is difficult to balance between the response of the pressure compensation valve and the response of the LS control of the hydraulic pump in controlling the swing drive pressure at the stage of transition from the acceleration to the steady rotation for the following reason. Can be mentioned.
[0008]
(1) At the time of turning start / acceleration, in order to maintain a constant flow rate, the pump LS control controls the discharge pressure of the hydraulic pump to be high according to the turning start pressure.
[0009]
(2) The pressure compensation valve operates in a direction to increase the passage flow rate that tends to decrease due to an increase in load pressure in order to keep the differential pressure before and after the throttle element of the direction switching valve constant.
[0010]
(3) Since the turning drive pressure decreases when turning reaches a steady speed, the pump LS control does not need to control the discharge pressure of the hydraulic pump higher at the time of start-up / acceleration, and operates in a direction to lower the discharge pressure of the hydraulic pump. .
[0011]
(4) The pressure compensation valve operates in a direction to reduce the passage flow rate that tends to increase due to a decrease in the swing drive pressure.
[0012]
Since the transition of (1) to (4) is steep, the turning operability is jerky (above (1)).
[0013]
Moreover, when the discharge flow rate of the hydraulic pump becomes a saturation state that does not satisfy the flow rate required by the plurality of directional control valves during the combined operation as described above, each target compensation differential pressure of the pressure compensation valve is changed according to the saturation state. The hydraulic pump discharge flow rate is reduced and redistributed to the flow rate ratio required by each actuator. Although this function reduces the speed of each actuator even during a combined operation, the actuator operates at a rate intended for the operation, so that the operational feeling is not impaired.
[0014]
However, this speed reduction occurs in the same manner with respect to the turning operation, and the turning speed is reduced in the same manner as other actuators in the combined operation including turning. This speed reduction causes a change in the turning speed when shifting from the combined turning operation to the turning single operation, or vice versa, giving the operator a sense of incongruity ((2) above).
[0015]
(2) JP 10-89304 A
In the hydraulic drive device described in Japanese Patent Laid-Open No. 10-89304, since the pressure compensation valve has a load-dependent characteristic, the target compensation differential pressure of the pressure compensation valve is determined according to the high load pressure of the swing motor when the swing is independently activated. When the state is lowered and the state shifts to a steady state, the target compensation differential pressure of the pressure compensation valve also returns to the original in accordance with the reduced load pressure of the swing motor, so that the turn can be started without a feeling of turning operability. However, when the discharge flow rate of the hydraulic pump is in a saturation state during the combined swing operation, it is disclosed in Japanese Patent Application Laid-Open No. 60-11706 that the discharge flow rate of the hydraulic pump is redistributed to the ratio of the flow rate required by each direction switching valve. It is the same as the hydraulic drive device described, and when changing from a combined turning operation to a turning single operation or vice versa, a change in the turning speed occurs, giving the operator a sense of incongruity (above (2)).
[0016]
In addition, since the pressure compensation valve has load dependent characteristics, the target compensation differential pressure of the pressure compensation valve of the swing section is reduced depending on the state of the discharge flow rate of the hydraulic pump at the time of combined turning startup. In addition, the target compensation differential pressure decreases due to the load-dependent characteristic in which the load pressure of the swing motor increases to the relief pressure, and the decrease in the target compensation differential pressure continues until the steady state shifts. As a result, the turning speed at the time of combined turning of the turning is extremely lower than that of other actuators, and the turning operability of the turning combined starting is impaired ((3) above).
[0017]
(3) Open center type three-pump system equipped with a hydraulic drive device and an actual machine described in JP-A-10-37907
In the hydraulic drive device described in Japanese Patent Laid-Open No. 10-37907, the turning operability is ensured in the LS system by configuring the turning control system with an open center type separate circuit. Further, even in an open center type three-pump system mounted on an actual machine, the turning control system is a separate circuit of the open center type, and the turning operability is ensured.
[0018]
In other words, in the case of the open center type, when the drive pressure rises at the time of turning start, the flow rate that returns to the tank through the center bypass oil passage increases, so the flow rate of the pressure oil that passes through the throttle of the direction switching valve of the turning section decreases. To do. For this reason, the flow rate of the pressure oil supplied to the turning motor is limited during startup and acceleration. When the turning speed reaches a steady speed, the drive pressure is not as high as that at the time of activation, so that the flow rate is not limited, and a flow rate equivalent to the opening of the throttle of the direction switching valve in the turning section is supplied to the turning motor. This makes it possible to smoothly start turning without causing a jerky feeling of operability when turning alone as in LS control.
[0019]
The above problem (2) occurs not only in the LS system but also in an open center type system. However, the hydraulic drive device described in Japanese Patent Laid-Open No. 10-37907 and an open center type three-pump system mounted on an actual machine. Then, by configuring the turning control system with a separate circuit of the open center type, independence of the turning control system is realized, and the turning speed does not change.
[0020]
However, in the open-center type three-pump system mounted on the hydraulic drive device and the actual machine described in Japanese Patent Laid-Open No. 10-37907, the turning control system must be configured in parallel with a circuit separate from other actuator systems. Accordingly, the cost is increased and the installation space is increased, and a hydraulic pump for the swing control system must be provided separately. In particular, in the system disclosed in Japanese Patent Laid-Open No. 10-37907, they are arranged in parallel. In order to balance the power with the LS system, a signal path is required and the circuit configuration becomes complicated ((4) above).
[0021]
It is an object of the present invention to provide a hydraulic drive system including a turning control system, which can be accelerated and shifted to a steady state without any jerky feeling of turning operability at the time of either turning alone or combined, and turning from turning alone operation. The change in turning speed when shifting to composite operation or vice versa is suppressed, and the turning speed is not extremely slow compared to other actuators at the start of composite operation, ensuring excellent turning operability and turning independence. At the same time, it is an object of the present invention to provide a hydraulic drive device that does not cause a problem of an increase in cost and space and a complicated circuit configuration due to the provision of another circuit.
[0022]
[Means for Solving the Problems]
(1) In order to achieve the above object, the present invention provides a hydraulic pump, a plurality of actuators including a swing motor driven by pressure oil discharged from the hydraulic pump, and the hydraulic pump to the plurality of actuators. A plurality of directional control valves for controlling the flow rate of the supplied pressure oil, a plurality of pressure compensating valves for controlling the differential pressure across the plurality of directional switching valves, and a discharge pressure of the hydraulic pump for the plurality of actuators In a hydraulic drive apparatus comprising a load sensing control pump control means for controlling a pump discharge flow rate so as to be higher than a maximum load pressure by a predetermined value, of the plurality of pressure compensation valves other than the swing section related to the swing motor The pressure compensation valve is used to compensate for the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the actuators. A first means for setting the pressure; a second means for setting the target compensation differential pressure; and a pressure compensation valve for at least the swing section among the plurality of pressure compensation valves. And the third means for reducing the target compensation differential pressure set by the second means to give the pressure compensation valve of the swing section a load dependent characteristic when the load pressure of the swing motor increases, and the swing Provided in the section pressure compensation valve, The lower limit of the target compensation differential pressure that is set by the second means and corrected by the third means is the lower limit of the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators. In order not to become smaller than the target compensation differential pressure set by one means, And a fourth means for setting a lower limit of the target compensation differential pressure set by the second means and corrected by the third means.
[0023]
In the present invention configured as described above, the third section is provided in the pressure compensation valve for the swing section so as to have a load-dependent characteristic, so that the pressure compensation for the swing section according to the change in the load pressure of the swing motor at the start of the swing. The valve finely adjusts the flow rate, and the swing motor accelerates smoothly and shifts to a steady state.
[0024]
Further, the second means for setting the target compensation differential pressure of the pressure compensation valve in the swing section is similar to the first means in that the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators is set as the target compensation differential. In this case, by providing the fourth means as described above, the fourth means reduces the target compensation differential pressure itself set by the second means and the third means. It functions as a lower limit setting means for both lowering of the target compensation differential pressure due to the load dependence characteristic given in (see (2) below). This causes the discharge flow rate of the hydraulic pump to saturate and the target compensation differential pressure of the pressure compensation valve in the swing section decreases, or the load pressure of the swing motor becomes high and the target compensation difference of the pressure compensation valve in the swing section. When the pressure tends to decrease due to the load-dependent characteristic or when they occur simultaneously, the fourth means limits the decrease of the target compensation differential pressure, and pressure oil is preferentially supplied to the swing motor. As a result, the turning speed change at the time of transition from the turning single operation to the turning combined operation or vice versa is suppressed, and the turning speed is not extremely slow compared to other actuators at the start of the composite, and excellent turning operation And turning independence can be secured.
[0025]
The second means for setting the target compensation differential pressure of the pressure compensation valve in the swing section is a means for setting a value that does not change depending on the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators as the target compensation differential pressure. In this case, the fourth means functions as a lower limit setting means with respect to a decrease in the target compensation differential pressure due to the load dependence characteristic given by the third means (see (3) below). . As a result, even when the discharge flow rate of the hydraulic pump becomes saturated, the target compensation differential pressure of the pressure compensation valve of the swing section does not decrease, and the load pressure of the swing motor becomes high and the target compensation of the pressure compensation valve of the swing section When the differential pressure is about to decrease due to the load-dependent characteristic, the fourth means limits the decrease of the target compensation differential pressure, so that whether the target compensation differential pressure is decreased due to saturation or load-dependent characteristics, either alone or simultaneously. The pressure oil is preferentially supplied to the swing motor. As a result, the turning speed change at the time of transition from the turning single operation to the turning combined operation or vice versa is suppressed, and the turning speed is not extremely slow compared to other actuators at the start of the composite, and excellent turning operation And turning independence can be secured.
[0026]
Furthermore, since the above function is achieved without providing a separate circuit, there is no problem of an increase in cost and space and a complicated circuit configuration.
[0027]
(2) In the above (1), preferably, the second means, like the first means, calculates the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators as the target compensation differential pressure. The fourth means is used for both the reduction of the target compensation differential pressure itself set by the second means and the reduction of the target compensation differential pressure due to the load dependence characteristic given by the third means. On the other hand, it functions as a lower limit setting means.
[0028]
As a result, as described in (1) above, when the discharge flow rate of the hydraulic pump is in a saturation state and the target compensation differential pressure of the pressure compensation valve of the swing section is about to decrease, or the load pressure of the swing motor becomes high. When the target compensation differential pressure of the pressure compensation valve of the swing section is about to decrease due to load dependent characteristics, or when they occur simultaneously, the fourth means limits the decrease of the target compensation differential pressure and gives priority to the swing motor. Pressure oil is supplied, and excellent turning operability and turning independence can be secured.
[0029]
(3) In the above (1), the second means sets the value that does not change due to the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators as the target compensation differential pressure. In this case, the fourth means functions as a lower limit setting means with respect to a decrease in the target compensation differential pressure due to the load dependence characteristic given by the third means.
[0030]
As a result, as described in (1) above, even if the discharge flow rate of the hydraulic pump is saturated, the target compensation differential pressure of the pressure compensation valve of the swing section does not decrease and the load pressure of the swing motor becomes high. When the target compensation differential pressure of the pressure compensation valve of the swing section is about to decrease due to the load dependent characteristic, the fourth means limits the decrease of the target compensation differential pressure, and the target compensation differential pressure due to saturation or the load dependent characteristic is limited. Regardless of whether the lowering occurs alone or simultaneously, pressure oil is preferentially supplied to the turning motor, and excellent turning operability and turning independence can be ensured.
[0031]
(4) Further, in the above (1) to (3), preferably, the fourth means is set by the second means, and when the target compensation differential pressure corrected by the third means reaches a predetermined value. The biasing means applies a biasing force in the opening direction to the spool of the pressure compensation valve of the swivel section.
[0032]
Thus, the fourth means sets the lower limit of the target compensation differential pressure without reducing the target compensation differential pressure of the pressure compensation valve of the turning section below the value corresponding to the biasing force applied by the biasing means.
[0033]
(5) In the above (4), preferably, when the target compensation differential pressure set by the second means and corrected by the third means reaches a predetermined value, the biasing means is a pressure of the swivel section. A lower limit setting spring that acts on the spool of the compensation valve and biases the spool in the opening direction.
[0034]
As a result, when the target compensation differential pressure of the pressure compensation valve of the swing section reaches a predetermined value, the biasing means applies a biasing force in the opening direction to the spool of the pressure compensation valve of the swing section.
[0035]
(6) In the above (1) and (2), preferably, the fourth means is always an auxiliary value for the target compensation differential pressure set by the second means and corrected by the third means. The direction switching valve of the swivel section has a meter-in variable throttle opening area other than the swivel section corresponding to the auxiliary compensation target compensation pressure added by the biasing means. It is comprised so that it may become smaller than the opening area of this direction switching valve.
[0036]
As a result, the fourth means limits the lowering of the target compensation differential pressure of the pressure compensation valve of the swing section by the auxiliary value added by the biasing means, and sets the lower limit of the target compensation differential pressure.
[0037]
(7) In the above (6), preferably, the urging means is a turning priority spring that always acts in the opening direction of the spool of the pressure compensation valve of the turning section.
[0038]
As a result, the biasing means always adds an auxiliary value to the target compensation differential pressure of the pressure compensation valve in the swing section.
[0039]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
[0040]
FIG. 1 shows a hydraulic drive apparatus according to a first embodiment of the present invention, and includes a plurality of actuators 2 including a hydraulic pump 1 and a swing motor 2 driven by pressure oil discharged from the hydraulic pump 1. 6, the difference between the front and rear of the plurality of directional control valves 7 to 11 of the closed center type and the plurality of directional control valves 7 to 11 that respectively control the flow rates of the pressure oil supplied from the hydraulic pump 1 to the plurality of actuators 2 to 6. A plurality of pressure compensation valves 12 to 16 that respectively control pressure, load check valves 17a to 17e that are disposed between the direction switching valves 7 to 11 and the pressure compensation valves 12 to 16 to prevent backflow of pressure oil, A pump controller 18 for load sensing control that controls the pump discharge flow rate so that the discharge pressure of the hydraulic pump 1 is higher than the maximum load pressure of the plurality of actuators 2 to 6 by a predetermined value. There. Overload relief valves 60 a and 60 b are provided in the actuator line of the swing motor 2. The other actuators 3 to 6 are also provided with the same overload relief valve, but the illustration is omitted.
[0041]
The plurality of directional control valves 7 to 11 are provided with self-load pressure detection lines 20 to 24. The maximum load pressure among the load pressures detected by the detection lines 20 to 24 is the signal lines 25 to 29, the shuttle valve. 30 to 33 and the signal lines 34 to 36 are detected and led to the signal line 37.
[0042]
The pump control device 18 includes a tilt control actuator 40 connected to a swash plate 1 a that is a capacity variable member of the hydraulic pump 1, a hydraulic chamber 40 a of the actuator 40, a discharge oil passage 1 b of the hydraulic pump 1, and a tank 19. A load sensing control valve (hereinafter referred to as an LS control valve) 41 for switching and controlling the connection. As the control pressure, the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the signal line 37 are opposed to the LS control valve. When the pump discharge pressure becomes higher than the total pressure of the maximum load pressure and the set value of the spring 41a (target LS differential pressure), the hydraulic chamber 40a of the actuator 40 is connected to the discharge oil passage 1b of the hydraulic pump 1, and the hydraulic chamber 40a The piston 40b is overcome by the force of the spring 40c by guiding the high pressure to the left side in the figure, and the tilt of the swash plate 1a is reduced to reduce the discharge flow rate of the hydraulic pump 1. Conversely, when the pump discharge pressure is lower than the total pressure of the maximum load pressure and the set value of the spring 41a (target LS differential pressure), the hydraulic chamber 40a of the actuator 40 is connected to the tank 19 and the hydraulic chamber 40a is depressurized. As a result, the piston 40b is moved rightward in the figure by the force of the spring 40c, and the displacement of the hydraulic pump 1 is increased by increasing the tilt of the swash plate 1a. By such operation of the LS control valve, the discharge flow rate of the hydraulic pump 1 is controlled so that the pump discharge pressure is higher than the maximum load pressure by the set value (target LS differential pressure) of the spring 41a.
[0043]
The pressure compensating valves 12 to 16 cause the pressure on the upstream side of the direction switching valves 7 to 11 to act in the closing direction, respectively, and the pressure (load) on the detection lines 20 to 24 that is the pressure on the downstream side of the direction switching valves 7 to 11. Pressure) is applied in the opening direction, and the maximum load pressure derived to the signal line 37 is applied in the closing direction, and the discharge pressure of the hydraulic pump 1 is applied in the opening direction, whereby the hydraulic pressure subjected to LS control as described above. The differential pressure between the discharge pressure of the pump 1 and the maximum load pressure (hereinafter referred to as LS control differential pressure as appropriate) is used as a target compensation differential pressure to control the differential pressure across the directional control valves 7 to 11.
[0044]
The upstream pressures of the respective direction switching valves 7 to 11 acting on the pressure compensating valves 12 to 16 are taken out by the signal lines 50a to 50e, and the detection lines 20 to 24 are the pressures downstream of the direction switching valves 7 to 11. The pressure (load pressure) of the hydraulic pump 1 is taken out by the signal lines 51a to 51e, the maximum load pressure of the signal line 37 is taken out by the signal lines 52 and 52a to 52e, and the discharge pressure of the hydraulic pump 1 is taken out by the signal lines 53 and 53a to 53e. It is taken out. In the pressure compensation valves 13 to 16, the maximum load pressure taken out by the signal lines 52b to 52e is loaded on the oil chambers 13a to 16a, and the discharge pressure of the hydraulic pump 1 taken out by the signal lines 53b to 53e is taken from the oil chambers 13b to 13b. 16b is set to set the target compensation differential pressure. The oil chamber for setting the target compensation differential pressure of the pressure compensation valve 12 will be described later.
[0045]
Further, the pressure compensation valve 12 causes the pressure on the upstream side of the direction switching valve 7 to act in the closing direction, and opens the pressure on the detection line 20 (the load pressure of the turning motor 2) that is the pressure on the downstream side of the direction switching valve 7. When the load pressure of the swing motor 2 rises when acting in the direction, it has a load dependent characteristic that reduces the target compensation differential pressure so as to limit the flow rate of the pressure oil passing through the direction switching valve 7. A lower limit setting spring 55 is provided on the opening direction acting side which is the setting side of the target compensation differential pressure. The lower limit setting spring 55 acts on the spool of the pressure compensation valve 12 only when the target compensation differential pressure of the pressure compensation valves 13 to 16 in the other sections becomes lower than the set value of the spring 55, and the target compensation differential pressure is reduced. The lower limit is set so as not to become smaller than the set value.
[0046]
The structure of the pressure compensation valve 12 is shown in FIG.
[0047]
In FIG. 2, the pressure compensation valve 12 has two bodies, a first body 301a and a second body 301b, and these bodies are assembled integrally by a method such as bolting (not shown) as appropriate. The first body 301a is provided with a small-diameter hole 321 and a medium-diameter hole 322 following the small-diameter hole 321. A first spool 311 having a diameter d1 is slidably fitted into the small-diameter hole 321 and the medium-diameter hole 322 is fitted. A second spool 312 having a diameter d3 (> d1) is slidably fitted. The second body 301b is provided with a large-diameter hole 323 following the medium-diameter hole 322 and a small-diameter hole 325 having the same diameter as the small-diameter hole 321 following the large-diameter hole 323. A third spool 310 is slidably fitted to 325, and the third spool 310 is slidably fitted to the large-diameter hole 323 and has first and second large-diameter portions 313 having a diameter d 2 (> d 3). , 314 and a small diameter portion 315 having a diameter d 1 slidably fitted into the small diameter hole 325.
[0048]
The end surface of the small diameter hole 321 is provided with a convex portion 321a, an oil chamber 331 is formed around the convex portion 321a, and the end surface of the first spool 311 is provided with a concave portion 311a for receiving the convex portion 321a. A weak spring 350 for holding the initial position is arranged between the end face of 321a and the bottom of the recess 311a to push the spools in the closing direction. The chamber in which the spring 350 is disposed communicates with the external oil chamber 331 through a passage 321b formed in the convex portion 321a.
[0049]
The lower limit setting spring 55 is disposed around the convex portion 321 of the oil chamber 331 and faces the end surface of the first spool 311. The lower limit setting spring 55 is only facing the end surface of the first spool 311 at the initial position shown in the drawing, and is separated from the end surface. Open There is no force pushing in the direction.
[0050]
The body 301a has a pump port 341 and a load pressure port 342, and the body 301b has a tank port 343, an outlet port 344, an inlet port 345, and a maximum load pressure port 346. The pump port 341 communicates with the discharge pressure signal line 53a of the hydraulic pump 1 and opens to the oil chamber 331. The load pressure port 342 communicates with the load pressure signal line 51a and has a small diameter hole 321 and a medium diameter hole 322. It opens to the oil chamber 332 formed in the connection part. Further, the tank port 343 opens to an oil chamber 333 provided in a large-diameter hole 323 that communicates with the tank 19 and surrounds the contact portion between the second spool 312 and the third spool 310, and the outlet port 344 has a load check Opened to an oil chamber 328 connected to the valve 17a and provided in the large-diameter hole 323 between the first and second spool large-diameter portions 313 and 314, the inlet port 345 communicates with the pump discharge oil passage 1b and The maximum load pressure port 346 communicates with the signal line 52a for the maximum load pressure and is connected to the signal line 52a for the maximum load pressure. An oil chamber 336 is provided in a portion of the large-diameter hole 323 where the continuous portion of the second large-diameter portion 314 and the small-diameter portion 315 is located.
[0051]
An oil chamber 334 communicating with the oil chamber 328 of the outlet port 344 is provided between the small diameter portion 315 and the small diameter hole end surface 330 via a pilot oil passage 50 a provided in the third spool 310.
[0052]
The first body 301a and the second body 301b are appropriately assembled together by a method such as bolting (not shown) as appropriate to form the body 301. At this time, the first body 301a-side medium-diameter hole 322 and the second body 301b are combined. Even if the side large-diameter hole 323 is misaligned, the second spool 312 and the third spool 310 are merely in contact with each other, so there is no problem in operation.
[0053]
With the above configuration, the pressure compensating valve 12 causes the outlet pressure (Pz) of the outlet port 344 in the closing direction to be applied to the pressure receiving area B1 of the end surface 340 of the small diameter portion 315 in the oil chamber 334 via the pilot oil passage 50a. The maximum load pressure (PLmax) of the port 346 is applied to the pressure receiving area B2 of the stepped portion obtained by subtracting the sectional area of the small diameter portion 315 from the sectional area of the second large diameter portion 314 in the oil chamber 336. In addition, the pressure compensation valve 12 opens the pump discharge pressure (Ps) through the pump port 341 to the pressure receiving area B1 of the end surface of the first spool 311 in the oil chamber 331, and the load pressure (PL) of the load pressure port 342. Are applied to the pressure receiving area B3 of the stepped portion obtained by subtracting the cross-sectional area B1 of the first spool 311 from the cross-sectional area of the second spool 312 in the oil chamber 332, respectively. The pressure receiving area of the stepped portion obtained by subtracting the cross-sectional area of the second spool 312 from the cross-sectional area of the first large-diameter portion 313 in the oil chamber 333 is that the oil chamber 33 communicates with the tank 19 through the tank port 343. The acting force for opening and closing each spool does not work.
[0054]
The pressure receiving area B2 and the pressure receiving area B1 of the first spool 311 are made substantially the same (B1 = B2), and the pressure receiving area B3 is made smaller than the pressure receiving area B1 (= B2) of the first spool (B1>). B3), which has a load-dependent characteristic that reduces the passage flow rate of the direction switching valve 7 leading to the swing motor 2 in accordance with an increase in the load pressure (PL) of the swing motor 2.
[0055]
In other words, considering the hydraulic balance of the first spool 311, the second spool 312, and the third spool 313, the pressure compensation valve 12 functions by balancing B1Pz-B3PLmax with respect to B1Ps-B2PLmax.
[0056]
B1Ps-B2PLmax = B1Pz-B3PL
From B1 = B2,
B1 (Ps-PLmax) = B2Pz-B3PL
Since Ps−PLmax is a differential pressure (LS control differential pressure) between the discharge pressure Ps of the hydraulic pump 1 subjected to LS control and the maximum load pressure PLmax, it is assumed that ΔPc.
B1ΔPc = B2Pz-B3PL (1)
When the differential pressure across the direction switching valve 7 is ΔP,
ΔP = Pz-PL
It becomes. Also, by transforming equation (1)
B1ΔPc + (B3−B2) PL = B2 (Pz−PL)
Therefore,
Figure 0003853123
Here, if B1 / B2 = α and B3 / B2 = β,
ΔP = Pz−PL = αΔPc− (1−β) PL (3)
That is, if B2 = B3 (if there is no area difference between B2 and B3),
ΔP = αΔPc
ΔP is determined only by ΔPc (LS control differential pressure). However, since there is an area difference when B2 ≠ B3, ΔP is affected by the load pressure PL due to the area difference, and ΔP is decreased as the load pressure PL increases. It has a load-dependent characteristic that reduces the passage flow rate of the direction switching valve 7.
[0057]
FIG. 3 shows the load-dependent characteristics of the pressure compensation valve 12. The horizontal axis in FIG. 3 is the load pressure, expressed as PL, and the vertical axis is the target compensation differential pressure, expressed as ΔPv. The dotted line shows the target compensation differential pressure of the pressure compensation valves 13 to 16 other than the section of the swing motor 2 (hereinafter referred to as the swing section) for reference. The pressure compensation valves 13 to 16 other than the swing section maintain the target compensation differential pressure ΔPv at the LS control differential pressure ΔPc even if the load pressure PL of the actuators 3 to 6 increases. No. 12, as the load pressure PL increases, the target compensation differential pressure ΔPv decreases as the load pressure PL increases.
[0058]
FIG. 4 shows a lower limit setting function of the target compensation differential pressure by the lower limit setting spring 55 when it is assumed that the pressure compensation valve 12 has no load dependent characteristics. The horizontal axis in FIG. 4 represents the sum of the flow rates required by the direction switching valve 7 and the other direction switching valves 8 to 11 (valve required flow rate), and is represented by Qr. This corresponds to the total amount of lever operation of an operation lever device (not shown) for switching the direction switching valves 7 to 11, that is, the total required flow rate of the swing motor 2 and its actuator. The vertical axis represents the target compensation differential pressure ΔPv set in the pressure compensation valve 12 and the other pressure compensation valves 13-16. The set differential pressure of the lower limit setting spring 55 (the lower limit value of the target compensation differential pressure) is Pb.
[0059]
At the time of the combined swing operation in which the swing motor 2 and other actuators are driven simultaneously, the sum Qr of the valve required flow rates of the direction switching valve 7 and the other direction switching valves 8 to 11 is smaller than the maximum discharge flow rate Qpmax of the hydraulic pump 1, When the discharge flow rate of the pump 1 is not in the saturation state, the target compensation differential pressure ΔPv of all the pressure compensation valves including the pressure compensation valve 12 is constant at the LS control differential pressure ΔPc.
[0060]
When the total required valve flow rate Qr exceeds the maximum discharge flow rate Qpmax of the hydraulic pump 1 and the discharge flow rate of the hydraulic pump 1 is in a saturation state, the LS control differential pressure ΔPc is set to the lower limit setting spring 55 of the pressure compensation valve 12 of the swing section. Until the pressure difference is reduced to the differential pressure Pb, the target compensation differential pressure ΔPv of all the pressure compensation valves decreases as the LS control differential pressure ΔPc decreases, and the LS control differential pressure ΔPc decreases to the set differential pressure Pb of the lower limit setting spring 55. After that, the target compensation differential pressure ΔPv of the pressure compensation valve 12 in the swing section is held at the set differential pressure Pb of the lower limit setting spring 55 and does not decrease below that, but the target compensation differential of the pressure compensation valves other than the swing section The pressure ΔPv continues to decrease as the LS control differential pressure ΔPc decreases.
[0061]
In the figure, the thick broken line is a change in the target compensation differential pressure ΔPv of the pressure compensation valves 13 to 16 other than the swivel section during the combined operation including the swivel section, and the thin broken line is a pressure during the combined operation not including the swivel section. This is a change in the target compensation differential pressure ΔPv of the compensation valves 13-16. The target compensation differential pressure ΔPv of the pressure compensation valves 13 to 16 other than the pivoting section in the combined operation including the pivoting section is such that the target compensation differential pressure ΔPv of the pressure compensation valve 12 of the pivoting section is greater than the set differential pressure Pb of the lower limit setting spring 55. Since it does not become small, the degree of decrease becomes larger than the target compensation differential pressure ΔPv of the pressure compensation valves 13 to 16 in the combined operation not including the turning section.
[0062]
The above hydraulic drive device is mounted on, for example, a hydraulic excavator. FIG. 5 shows the external appearance of the hydraulic excavator. In FIG. 5, the excavator includes a lower traveling body 200, an upper swing body 201, and a front work machine 202, and the upper swing body 201 can swing around the axis O on the lower travel body 200. Can move up and down at the front of the upper swing body 201. The front work machine 202 has an articulated structure including a boom 203, an arm 204, and a bucket 205. The boom 203 includes a boom cylinder 206, the arm 204 includes an arm cylinder 207, and the bucket 205 includes a bucket cylinder 208. The inside is driven to rotate. The turning motor 2 shown in FIG. 1 is an actuator that drives the upper turning body 202 to turn on the lower traveling body 200, and three of the actuators 3 to 6 are used as a boom cylinder 206, an arm cylinder 207, and a bucket cylinder 208.
[0063]
In the above, the oil chambers 13a to 16a and 13b to 16b connected to the signal lines 52b to 52e and 53b to 53e of the pressure compensation valves 13 to 16 are the turning sections related to the turning motor 2 among the plurality of pressure compensation valves 12 to 16. Pressure compensation valves 13 to 16 other than the above are configured, and a first means for setting a differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the plurality of actuators 2 to 6 as a target compensation differential pressure is configured. The oil chamber 336 (pressure receiving area B2 = B1) and the oil chamber 331 (pressure receiving area B1) connected to the signal lines 52a and 53a of the valve 12 are provided in the pressure compensating valve 12 of the swivel section, and set the target compensation differential pressure. The oil chamber 334 (pressure receiving area B1> B3) and the oil chamber 332 (pressure receiving area B3) constituting the second means and connected to the signal lines 50a and 51a of the pressure compensation valve 12 are composed of a plurality of pressures. Among the compensation valves 12-16, provided at least in the pressure compensation valve 12 of the swing section, when the load pressure of the swing motor 2 rises, the target compensation differential pressure set by the second means is reduced, and the pressure of the swing section is increased. A third means for imparting load dependent characteristics to the compensation valve 12 is configured, and a lower limit setting spring 55 of the pressure compensation valve 12 is provided in the pressure compensation valve 12 of the swing section, set by the second means, and the third means. A fourth means for setting a lower limit of the target compensation differential pressure corrected by the means is configured.
[0064]
In the present embodiment, the second means (oil chambers 331 and 336) is similar to the first means (oil chambers 13a to 16a and 13b to 16b), and the discharge pressure of the hydraulic pump 1 and a plurality of actuators 2 to 6 are used. The above four means (lower limit setting spring 55) are the target compensation differential pressure itself set by the second means (oil chambers 331, 336). It functions as a lower limit setting means for both the decrease of the target compensation differential pressure due to the load dependence characteristic given by the third means (oil chambers 332, 334).
[0065]
Further, the fourth means (lower limit setting spring 55) is set by the second means (oil chambers 331 and 336), and the target compensation differential pressure corrected by the third means (oil chambers 332 and 334) becomes a predetermined value. When it reaches, the biasing means applies a biasing force in the opening direction to the spool 311 of the pressure compensation valve 12 of the swing section.
[0066]
The operation of the present embodiment configured as described above will be described.
1. When turning alone
FIG. 6 is a time chart showing the behavior of the pressure compensating valve 12 for turning during the turning single operation in which the turning direction switching valve 7 is operated to drive the turning motor 2 alone.
[0067]
When the single swing operation is started, there is an increase in the load pressure peculiar to the inertia load of the upper swing body 201. This increase in load pressure is limited by a safety valve that is an overload relief valve 60a or 60b provided in the swing motor 2. In this state, the pressure oil supplied to the turning motor 2 is discharged from the safety valve 60a or 60b to the tank.
[0068]
In the conventional general pressure compensation valve, the acceleration feeling of the upper-part turning body 201 which is an inertial load is adjusted by releasing the pressure oil from the safety valve. However, in this case, since the consumption flow rate of the turning motor at the time of startup is small, most of the pressure oil is discharged to the tank, resulting in energy loss. Further, it is difficult to balance the LS control of the hydraulic pump and the flow rate compensation function of the pressure compensation valve, and the turning operability becomes jerky.
[0069]
On the other hand, in this embodiment, since the pressure compensation valve 12 in the swing section has the load-dependent characteristics as described above, such a problem does not occur.
[0070]
First, in a state before activation in which the operation lever of the turning operation lever device is not operated, the target compensation differential pressure ΔPv of the pressure compensation valve 12 is controlled to the LS control differential pressure ΔPc (t0 to t1).
[0071]
When the swing motor 2 is started by operating the operation lever, the load pressure PL increases due to the inertia load simultaneously with the start (t1).
[0072]
Due to the load-dependent characteristics of the pressure compensation valve 12, the target compensation differential pressure ΔPv decreases from the LS control differential pressure ΔPc and stops decreasing at the set differential pressure Pb of the lower limit setting spring 55 (t1). The supply flow rate Qa to the swing motor 2 is controlled to a flow rate corresponding to the set differential pressure Pb of the spring 55. When the lower limit setting spring 55 is not provided, the target compensation differential pressure ΔPv is lowered to a pressure lower than Pb (not 0).
[0073]
When the upper swing body 201 starts rotating and the swing speed increases, the consumption flow rate of the swing motor 2 and the supply flow rate Qa to the swing motor 2 are balanced, and the load pressure gradually decreases. As a result, the target compensation differential pressure ΔPv of the pressure compensation valve 12 also increases (t2).
[0074]
When the consumption flow rate of the swing motor 2 and the supply flow rate Qa are not balanced, the load pressure PL is increased or decreased and fed back to the pressure compensation valve 12 of the swing section. Due to the load pressure-dependent characteristics of the pressure compensation valve 12, if the supply flow rate Qa is too large, the load pressure PL increases, and as a result, the supply flow rate Qa is limited by the pressure compensation valve 12. Conversely, when the supply flow rate Qa is insufficient, the load pressure PL decreases and the supply flow rate Qa is increased by the pressure compensation valve 12. By fine adjustment of the pressure compensation valve 12, the swing motor 2 is slowly accelerated without causing hunting that occurs in the conventional LS control.
[0075]
When the original supply flow rate is reached, a steady state is reached (t3), and the load pressure PL becomes a pressure corresponding to the rotational resistance.
2. Activation of other actuators during steady rotation of rotation
FIG. 7 is a time chart showing the behavior of the pressure compensation valve in each section when another actuator, for example, a boom cylinder, is activated and performs a combined operation while the vehicle is normally rotated by turning alone. It is assumed that the boom cylinder is the actuator 3.
[0076]
At the time of single turn steady rotation, the load pressure PL of the swing motor 2 is lowered to the pressure required for steady rotation, and the target compensation differential pressure ΔPv of the pressure compensation valve 12 is controlled to be approximately the LS control differential pressure ΔPc (t0˜ t1).
[0077]
When the operation lever of the boom operation lever device is additionally operated, the flow rate required by the swing motor 2 and the boom cylinder 3 exceeds the maximum discharge flow rate that can be supplied by the hydraulic pump 1, and when saturation occurs, the required flow rate The target compensation differential pressure ΔPv of each pressure compensation valve 12, 13 decreases due to a decrease in the LS control differential pressure ΔPc proportional to the supply shortage with respect to Qr, and flow redistribution occurs (t 1).
[0078]
Here, when the degree of saturation is large, the target compensation differential pressure ΔPv greatly decreases, but the decrease in the target compensation differential pressure ΔPv of the pressure compensation valve 12 in the swing section is limited by the set differential pressure Pb of the lower limit setting spring 55. The For this reason, the target compensation differential pressure ΔPv of the pressure compensation valve 13 in the boom section is further lowered by the amount by which the decrease in the target compensation differential pressure ΔPv on the turning side is limited.
[0079]
As a result, it is possible to preferentially supply pressure oil to the swing motor 2 to some extent during the combined operation including turning. With this function, it is possible to realize independent operability with respect to the other actuators of the swing motor 2 in the saturation state, and it is possible to suppress the swing speed change during the combined operation and to ensure the swing operability.
[0080]
As a comparative example, in a combined operation that does not include turning, the target compensation differential pressure ΔPv decreases to the same value due to a decrease in the LS control differential pressure ΔPc due to saturation, and the supply flow rate Qa also decreases to the same value (direction switching related to the combined operation). The valve opening area is assumed to be the same). The same applies to the combined operation including turning when the pressure compensating valve 12 of the turning section does not have the lower limit setting spring 55 (in the case of Japanese Patent Laid-Open No. 10-89304). Even in this case, the target compensation differential pressure ΔPv and the supply flow rate Qa of the swing section are prevented from being lowered by ΔΔPv1 and ΔQa1, pressure oil is preferentially supplied to the swing motor 2, and the change in the speed of the swing during the combined operation can be suppressed.
[0081]
FIG. 8 shows a case where the degree of saturation of the discharge flow rate of the hydraulic pump 1 in the combined operation is small.
[0082]
When the degree of saturation is small, the decrease in the target compensation differential pressure ΔPv stays above the set differential pressure Pb of the lower limit setting spring 55. In this case, both the swing and the boom are reduced to the same target compensation differential pressure ΔPv and the flow rate Qa (assuming that the opening areas of the direction switching valves 7 and 8 in the swing and boom sections are the same).
[0083]
Thus, by setting the lower limit setting spring 55, it becomes possible to set the priority of turning by the degree of saturation.
3. Swing and simultaneous activation with other actuators
FIG. 9 is a time chart showing the behavior of the pressure compensation valve in each section during a combined operation in which another actuator, for example, a boom cylinder is activated at the same time when the turning is activated. Also in this case, it is assumed that the boom cylinder is the actuator 3.
[0084]
First, in a state before activation in which the operation levers of the turning and boom operation lever devices are not operated, the target compensation differential pressure ΔPv of the pressure compensation valves 12 and 13 is controlled to the LS control differential pressure ΔPc (t0 to t1). ).
[0085]
When the swing and boom control levers are simultaneously operated to simultaneously start the swing motor 2 and the boom cylinder 3, the required flow rate when the swing and the boom are combined exceeds the maximum discharge flow rate of the hydraulic pump 1 and a saturation occurs. A decrease in the LS control differential pressure ΔPc proportional to the supply shortage with respect to the flow rate Qr causes the target compensation differential pressure ΔPv of each pressure compensation valve 12-16 to drop, and flow redistribution occurs (t1).
[0086]
Also in this case, by fine adjustment based on the load-dependent characteristics of the pressure compensation valve 12 in the swing section, the swing motor 2 is gradually accelerated without causing hunting that occurs in the conventional LS control.
[0087]
Further, when the degree of saturation is large, the target compensation differential pressure ΔPv is greatly reduced. Further, with respect to the pressure compensation valve 12 in the swing section, the load pressure PL of the swing motor 2 increases due to the inertia load simultaneously with the start of the swing motor 2, so that the target compensation differential pressure ΔPv also depends on the load dependent characteristics of the pressure compensation valve 12. There is a decline. The decrease in the target compensation differential pressure ΔPv of the pressure compensation valve 12 is limited by the set differential pressure Pb of the lower limit setting spring 55. For this reason, the target compensation differential pressure ΔPv of the pressure compensation valve 13 in the boom section is further lowered by the amount by which the decrease in the target compensation differential pressure ΔPv on the turning side is limited.
[0088]
As a result, the discharge flow rate of the hydraulic pump 1 is preferentially supplied to the turning motor 2 to some extent, and the turning operability can be maintained without the turning speed being extremely slow compared to the boom cylinder 3. The
[0089]
As a comparative example, in a combined operation that does not include turning, as shown by a broken line in FIG. 9, the target compensation differential pressure ΔPv decreases to the same value due to the decrease in the LS control differential pressure ΔPc due to saturation, and the supply flow rate Qa also reaches the same value. (It is assumed that the opening area of the directional control valve related to the combined operation is the same).
[0090]
In the combined operation including turning when the pressure compensating valve 12 of the turning section does not have the lower limit setting spring 55 (Japanese Patent Laid-Open No. 10-89304), the LS control differential pressure ΔPc is reduced due to saturation and the load dependence of the pressure compensating valve 12 Depending on the characteristics, the target compensation differential pressure ΔPv decreases extremely as shown by the two-dot chain line in FIG. 9, and the supply flow rate Qa also decreases extremely. In the present embodiment, the decrease in the target compensation differential pressure ΔPv of the pressure compensation valve 12 is limited by the set differential pressure Pb of the lower limit setting spring 55. For this reason, compared with the case where the spring 55 is not provided, a decrease in the target compensation differential pressure ΔPv and the supply flow rate Qa of the turning section can be suppressed by ΔΔPv2, ΔQa2. With this function, it is possible to maintain the turning operability during the combined operation without causing the turning speed to be extremely slow compared to other actuators.
[0091]
FIG. 10 shows a case where the degree of saturation of the discharge flow rate of the hydraulic pump 1 in the combined operation is small.
When the degree of saturation is small, the decrease in the target compensation differential pressure ΔPv of the pressure compensation valve 13 in the boom section remains above the set differential pressure Pb of the lower limit setting spring 55 of the pressure compensation valve 12 in the swing section. Due to the load dependency of the pressure compensation valve 12 in the swing section, the target compensation differential pressure ΔPv in the swing section decreases to the set differential pressure Pb of the lower limit setting spring 55.
[0092]
As the turning speed increases, the load pressure of the turning motor 2 decreases, and the target compensation differential pressure ΔPv of the pressure compensation valve 12 in the turning section increases. Eventually, the target compensation differential pressure ΔPv and the supply flow rate Qa are the same for both the turning and boom sections (assuming that the opening area of the direction switching valve of the turning and boom sections is the same) (t4).
[0093]
When the pressure compensation valve 12 in the swing section does not have the lower limit setting spring 55 (Japanese Patent Laid-Open No. 10-89304), the target compensation differential pressure of the pressure compensation valve 12 in the swing section is shown by a two-dot chain line in FIG. ΔPv decreases to a pressure lower than Pb, and the supply flow rate Qa to the swing motor 2 also decreases significantly immediately after startup. By providing the lower limit setting spring 55, a decrease in the target compensation differential pressure ΔPv and the supply flow rate Qa of the turning section can be suppressed by ΔΔPv3 and ΔQa3 immediately after startup as compared with that case. Therefore, also in this case, the turning operability can be maintained without extremely reducing the turning speed as compared with other actuators.
[0094]
As described above, according to the present embodiment, due to the load-dependent characteristics of the pressure compensation valve 12 in the swing section, even when turning alone or in combination, there is no jerky feeling of turning operability and the state shifts to a steady state. it can. Further, the lower limit setting spring 55 is provided in the pressure compensation valve 12 of the swing section so that the pressure oil is preferentially supplied to the swing motor 2 when the discharge flow rate of the hydraulic pump 1 is saturated. The change in turning speed is suppressed at the time of transition to the same, and the same is true at the time of transition from reverse swivel compound to swivel single operation, and at the start of swivel compound, the swivel speed is accelerated without being extremely slow compared to other actuators. It is possible to ensure excellent turning operability and turning independence. In addition, since the above function is achieved without providing a separate circuit, there is no problem of an increase in cost and space and a complicated circuit configuration.
[0095]
A second embodiment of the present invention will be described with reference to FIGS. In the figure, member loads equivalent to those shown in FIGS. 1 and 2 are given the same reference numerals. In this embodiment, the turning priority spring is always applied to the spool of the pressure compensation valve.
[0096]
In FIG. 11, the pressure compensating valves 13 to 16 other than the swivel section are the same as those in the first embodiment.
[0097]
The pressure compensation valve 12A in the swing section applies the pressure on the upstream side of the direction switching valve 7A in the closing direction, and opens the pressure (load pressure) of the detection lines 20 to 24, which is the pressure on the downstream side of the direction switching valve 7A. And the maximum load pressure derived to the signal line 37 is applied in the closing direction, and the discharge pressure of the hydraulic pump 1 is applied in the opening direction, whereby the LS control differential pressure (the discharge of the LS-controlled hydraulic pump 1 is discharged). The differential pressure between the directional control valve 7A and the directional switching valve 7A is controlled by using the differential pressure between the pressure and the maximum load pressure as a target compensation differential pressure, and when the load pressure of the swing motor 2 rises, it passes through the directional switching valve 7A. The pressure compensation valve 12 is configured to have a load-dependent characteristic that reduces the target compensation differential pressure so as to limit the flow rate of the pressure oil.
[0098]
The pressure compensation valve 12A has a turning priority spring 55A on the opening direction acting side, which is the setting side of the target compensation differential pressure, and this turning priority spring 55A is always connected to the pressure compensation valve 12A during the operation of the pressure compensation valve 12A. A certain auxiliary target compensation differential pressure for turning priority that acts on the spool and is added to the target compensation differential pressure by the LS control differential pressure is set. That is, the target compensation differential pressure of the pressure compensation valve 12A is larger than the pressure compensation valves 13 to 16 other than the turning section by a set amount by the turning priority spring 55A.
[0099]
Further, the direction switching valve 7A of the swing section obtains a flow rate characteristic as designed when the discharge flow rate of the hydraulic pump 1 is not in a saturation state corresponding to the setting of the larger target compensation differential pressure of the pressure compensation valve 12A. Therefore, the aperture areas of the meter-in variable apertures 57a and 57b are set smaller than usual.
[0100]
FIG. 12 shows the relationship. In the figure, M1 is a change in the opening area of the meter-in variable throttles 57a and 57b with respect to the spool stroke of the direction switching valve 7A (opening area characteristic), and M2 is a rated condition in which the swing priority spring 55A is not used for the pressure compensation valve. Is a change (opening area characteristic) of the opening area of the meter-in variable throttle with respect to the spool stroke of the direction switching valve (for example, the direction switching valve 7 in the first embodiment shown in FIG. 1). M1 has a larger opening area for the same spool stroke than M2 small It is set to become.
[0101]
The structure of the pressure compensation valve 12A is shown in FIG. In FIG. 13, a small diameter hole 321 having an end surface 320 is formed in the first body 301 a, and a first spool 311 that fits into the small diameter hole 321 in an oil chamber 331 </ b> A of the end surface 320 of the small diameter hole 321. Between the end surface 320 of the small diameter hole 321, the above-mentioned turning priority spring 55A that pushes the first spool 311, the second spool 312 and the third spool 310 in the closing direction is arranged. The relationship between the pressure receiving areas B1, B3, B1, and B2 in the oil chambers 331A, 332, 334, and 336 is the pressure receiving areas B1, B3, and B3 in the oil chambers 331, 332, 334, and 336 shown in FIG. This is the same as the relationship between B1 and B2. The other configuration of the pressure compensating valve 12A is the same as that of the first embodiment shown in FIG.
[0102]
The operating principle of the turning priority spring 55A in the pressure compensation valve 12A will be described.
[0103]
The lower limit setting spring 55 in the pressure compensation valve 12 of the first embodiment sets a lower limit for the target compensation differential pressure so that the target compensation differential pressure does not become lower than a predetermined value. Assuming that the lower limit value of the target compensation differential pressure is the aforementioned Pb, in this embodiment, the swing priority spring 55A is always applied to the spool, and the target compensation differential pressure corresponding to the lower limit value Pb is set as the target compensation by the LS control differential pressure. Set to be added to the differential pressure. As a result, the target compensation differential pressure of the pressure compensation valve 12A becomes larger by Pb than the other pressure compensation valves 13-16. That is,
Target compensation differential pressure of pressure compensation valves 13 to 16: Ps-PLmax
Target compensation differential pressure of pressure compensation valve 12A: Ps-PLmax + Pb
When the target compensation differential pressure of the pressure compensation valve 12A is set in this way, if the opening area of the meter-in variable throttle of the direction switching valve of the swing section is set to the same size as before, only the swing motor 2 is provided for Pb. A lot of flow will flow. Accordingly, it is necessary to reduce the opening area of the meter-in variable throttle of the direction switching valve of the swing section by Pb so that the same flow rate as before flows through the swing motor 2.
[0104]
That is, assuming that the opening area of the turning direction switching valve at the target compensation differential pressure of the original rated condition is As and the opening area of the meter-in variable throttle of the direction switching valve 7A is Aso,
Aso = As√ ((Ps−PLmax) / (Ps−PLmax + Pb))
It becomes.
[0105]
Changes in the supply flow rate to the swing motor 2 during saturation when such a pressure compensation valve 12A and the direction switching valve 7A are used will be compared with other actuators. The opening area of the direction switching valve related to the other actuator is set to As which is the same as the opening area of the direction switching valve for turning at the target compensation differential pressure under the rated conditions, the supply flow rate to the turning motor 2 is set to Qa, When the supply flow rate is Qb, Qa and Qb can be expressed as follows.
[0106]
Figure 0003853123
Here, As√ ((Ps−PLmax) / (Ps−PLmax + Pb)) is a value (constant number) under rated conditions.
[0107]
Set the rating conditions as follows.
[0108]
Figure 0003853123
These values are substituted into the above formulas Qb and Qa.
[0109]
Qb = 21.94√ΔPc
Qa = 21.94 × 0.91√ (ΔPc + Pb)
FIG. 14 shows a comparison of the relationship between the above Qa, Qb and the LS control differential pressure ΔPc. As can be seen from this figure, when the LS control differential pressure ΔPc is 15 kgf / cm 2 or less, that is, when the discharge flow rate of the hydraulic pump 1 is in a saturation state where the required flow rate is not reached, the supply flow rate Qa of the swing motor 2 is an actuator other than swing. Therefore, the pressure oil is preferentially supplied to the turning motor 2. Further, the degree of priority (flow rate difference) increases as the LS control differential pressure ΔPc decreases.
[0110]
In the above, the oil chambers 13a to 16a and 13b to 16b connected to the signal lines 52b to 52e and 53b to 53e of the pressure compensation valves 13 to 16 are the turning sections related to the turning motor 2 among the plurality of pressure compensation valves 12 to 16. Pressure compensation valves 13 to 16 other than the above are configured, and a first means for setting a differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the plurality of actuators 2 to 6 as a target compensation differential pressure is configured. An oil chamber 336 (pressure receiving area B2 = B1) and an oil chamber 331A (pressure receiving area B1) connected to the signal lines 52a and 53a of the valve 12A are provided in the pressure compensation valve 12 of the swing section, and set the target compensation differential pressure. There are a plurality of oil chambers 334 (pressure receiving area B1> B3) and oil chambers 332 (pressure receiving area B3) connected to the signal lines 50a and 51a of the pressure compensation valve 12A. Among the pressure compensation valves 12 to 16, provided at least in the pressure compensation valve 12A of the swing section, when the load pressure of the swing motor 2 increases, the target compensation differential pressure set by the second means is reduced, and The pressure compensation valve 12A constitutes a third means for giving a load-dependent characteristic, and the swing priority spring 55A of the pressure compensation valve 12A is provided in the pressure compensation valve 12A of the swing section, set by the second means, The fourth means for setting the lower limit of the target compensation differential pressure corrected by the three means is configured.
[0111]
Further, in the present embodiment, the second means (oil chambers 331A, 336) is similar to the first means (oil chambers 13a-16a, 13b-16b) and the discharge pressure of the hydraulic pump 1 and the plurality of actuators 2-6. The four means (swing priority springs 55) are the target compensation differential pressure itself set by the second means (oil chambers 331A, 336). It functions as a lower limit setting means for both the decrease of the target compensation differential pressure due to the load dependence characteristic given by the third means (oil chambers 332, 334).
[0112]
Further, the fourth means (the turning priority spring 55) is always supplementary to the target compensation differential pressure set by the second means (oil chambers 331A, 336) and corrected by the third means (oil chambers 332, 334). The direction switching valve 7A of the swing section has an opening area of the meter-in variable throttles 57a and 57b corresponding to the target compensation pressure corresponding to the auxiliary value added by the biasing means. Only, it is comprised so that it may become smaller than the opening area of direction switching valves 8-11 other than a turning section.
[0113]
Therefore, also in this embodiment, due to the load-dependent characteristics of the pressure compensation valve 12A of the turning section, it is possible to accelerate and shift to a steady state without any jerky feeling of the turning operability at the time of either turning alone or complex starting. In addition, a swing priority spring 55A is provided in the pressure compensation valve 12A of the swing section so that pressure oil is preferentially supplied to the swing motor 2 when the discharge flow rate of the hydraulic pump 1 is saturated. The change in turning speed is suppressed at the time of transition to the same, and the same is true at the time of transition from reverse swivel compound to swivel single operation, and at the start of swivel compound, the swivel speed is accelerated without being extremely slow compared to other actuators. It is possible to ensure excellent turning operability and turning independence. In addition, since the above function is achieved without providing a separate circuit, there is no problem of an increase in cost and space and a complicated circuit configuration.
[0114]
A third embodiment of the present invention will be described with reference to FIGS. In the figure, member loads equivalent to those shown in FIGS. 1 and 2 are given the same reference numerals. In the present embodiment, the turning priority is given to the pressure compensating valve of the turning section without setting the target compensation differential pressure by the LS control differential pressure.
[0115]
In FIG. 15, the pressure compensation valves 13 to 16 other than the swivel section are the same as those in the first embodiment.
[0116]
Further, the pressure compensation valve 12B of the swing section applies the pressure upstream of the direction switching valve 7 in the closing direction, and the pressure of the detection line 20 (load pressure of the swing motor 2) that is the pressure downstream of the direction switching valve 7. ) In the opening direction, when the load pressure of the swing motor 2 rises, the load dependent characteristic is such that the target compensation differential pressure is reduced so as to limit the flow rate of the pressure oil passing through the pressure compensation valve 12B. This point is the same as the pressure compensation valve 12 of the first embodiment.
[0117]
The pressure compensation valve 12B has means for setting a normal target compensation differential pressure, for example, a setup spring 60 on the opening direction acting side, which is the target compensation differential pressure setting side. The target compensation differential pressure is set to the same magnitude as the target compensation differential pressure due to the LS control differential pressure when the discharge flow rate is not in the saturation state. That is. When the discharge flow rate of the hydraulic pump 1 is in a saturation state, the pressure compensation valves 13 to 16 other than the swing section for setting the target compensation differential pressure based on the LS control differential pressure have a target compensation differential pressure that decreases according to the degree of saturation. On the other hand, in the pressure compensation valve 12B of the swing section, the target compensation differential pressure set by the setting spring 60 is substantially unchanged even when in the saturation state, and this target compensation differential pressure changes depending on the load dependent characteristics. .
[0118]
Further, the pressure compensation valve 12B is provided with a lower limit setting spring 55 for setting the lower limit of the target compensation differential pressure of the pressure compensation valve 12B, as in the first embodiment.
[0119]
The structure of the pressure compensation valve 12B is shown in FIG. In FIG. 16, the oil chambers 331, 336 in the first embodiment shown in FIG. 2 are replaced with oil chambers 331B, 336B, respectively, and these oil chambers 331B, 336B communicate with the tank via the tank ports 341B, 346B, respectively. The pressure receiving area B1 of the oil chamber 331B provided by the first spool 311 and the pressure receiving area B2 of the oil chamber 336B provided by the step portion between the second large diameter portion 314 and the small diameter portion 325 of the third spool 310 are respectively An oil pressure is not applied to the first spool 311 and the third spool 310. In addition, a spring 60 for setting the above-described target compensation differential pressure is disposed in the recess 311a formed on the end surface of the first spool 311 in place of the weak spring 350 for maintaining the initial position. The relationship between the pressure receiving areas B3 and B1 located in the oil chambers 332 and 334 is the same as that in the first embodiment (B1> B3), so that the swing motor 2 is increased according to the increase in the load pressure (PL) of the swing motor 2. It has a load-dependent characteristic that reduces the flow rate of the direction switching valve 7 that leads to.
[0120]
In the above, the oil chambers 13a to 16a and 13b to 16b connected to the signal lines 52b to 52e and 53b to 53e of the pressure compensation valves 13 to 16 are the turning sections related to the turning motor 2 among the plurality of pressure compensation valves 12 to 16. Pressure compensation valves 13 to 16 other than the above are configured, and a first means for setting a differential pressure between the discharge pressure of the hydraulic pump 1 and the maximum load pressure of the plurality of actuators 2 to 6 as a target compensation differential pressure is configured. The setting spring 60 of the valve 12B is provided in the pressure compensation valve 12B of the swing section, constitutes a second means for setting the target compensation differential pressure, and an oil chamber 334 (which is connected to the signal lines 50a and 51a of the pressure compensation valve 12B) The pressure receiving area B1> B3) and the oil chamber 332 (pressure receiving area B3) are provided at least in the pressure compensating valve 12B of the swing section among the plurality of pressure compensating valves 12-16. When the load pressure of the swing motor 2 increases, the target compensation differential pressure set by the second means is reduced, and a third means for providing the load compensation characteristic to the pressure compensation valve 12B of the swing section is constituted. The lower limit setting spring 55 of the valve 12 is provided in the pressure compensating valve 12 of the swing section, and constitutes fourth means for setting the lower limit of the target compensation differential pressure that is set by the second means and corrected by the third means. To do.
[0121]
In the present embodiment, the second means (setting spring 60) sets a value that does not change due to the differential pressure between the discharge pressure of the hydraulic pump 11 and the maximum load pressure of the plurality of actuators 2 to 6 as the target compensation differential pressure. The fourth means (lower limit setting spring 55) functions as a lower limit setting means for lowering of the target compensation differential pressure due to the load-dependent characteristics given by the third means (oil chambers 332, 334).
[0122]
Further, the fourth means (lower limit setting spring 55) is set by the second means (setting spring 60), and when the target compensation differential pressure corrected by the third means (oil chambers 332, 334) reaches a predetermined value. The biasing means applies a biasing force in the opening direction to the spool 311 of the pressure compensation valve 12B of the swing section.
[0123]
In the present embodiment configured as described above, the setting spring 60 sets the target compensation differential pressure having the same magnitude as the target compensation differential pressure due to the LS control differential pressure when the discharge flow rate of the hydraulic pump 1 is not in the saturation state. Therefore, before the discharge flow rate of the hydraulic pump 1 is saturated, the target is to distribute the discharge flow rate of the hydraulic pump 1 at the ratio of the required flow rates of the plurality of actuators as in the first embodiment. When the compensation differential pressure is set and the target compensation differential pressure is corrected by the load dependent characteristic of the pressure compensation valve 12B of the swing section, while the discharge flow rate of the hydraulic pump 1 is in the saturation state, the pressure compensation valve other than the swing section While the target compensation differential pressures 13 to 16 decrease as the LS control differential pressure decreases, the pressure compensation valve in the swing section decreases. The target compensation differential pressure due to the setting spring 60 of 2B does not change depending on the degree of saturation, the target compensation differential pressure of the pressure compensation valve 12B changes only depending on the load dependent characteristic, and the target compensation differential pressure decreases due to this load dependent characteristic. On the other hand, the lower limit setting spring 55 functions, and in this case as well, the pressure oil is preferentially supplied to the swing motor 2 as in the first and second embodiments.
[0124]
Therefore, according to this embodiment as well, due to the load-dependent characteristics of the pressure compensation valve 12B of the turning section, the turning operation can be accelerated and transitioned to a steady state without any jerky feeling of turning operability even when turning alone or in combination. Further, a lower limit setting spring 55 and a setting spring 60 are provided in the pressure compensation valve 12B of the swing section, and the pressure is preferentially applied to the swing motor 2 when saturation of the discharge flow rate of the hydraulic pump 1 and when the target compensation differential pressure decreases due to load dependent characteristics. Since oil is supplied, the change in the turning speed during the transition from turning single operation to turning combined operation is suppressed, and the same is true when switching from reverse turning composite operation to turning single operation. Compared with other actuators, the turning speed can be accelerated without being extremely slow, and excellent turning operability and turning independence can be secured. In addition, since the above function is achieved without providing a separate circuit, there is no problem of an increase in cost and space and a complicated circuit configuration.
[0125]
In the above embodiment, an example using a before-orifice type pressure compensation valve located on the upstream side of the direction switching valve is shown, but an after-orifice type pressure compensation valve located on the downstream side of the direction switching valve is used. However, it is possible to construct a system having the same effect.
[0126]
In the above embodiment, the lower limit setting spring 55, the rotation priority spring 55A, and the setting spring 60 are provided as means for controlling the target compensation differential pressure so as to give priority to the pressure compensation valve of the swing section. Similar to the oil chamber to which the upstream and downstream pressures are guided, an oil chamber for guiding the control pressure may be provided to apply a hydraulic control force. In this case, it is possible to perform more complicated and advantageous control by changing the control pressure according to the purpose.
[0127]
Further, in the above embodiment, in order to set the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators as the target compensation differential pressure, the pump discharge pressure and the maximum load pressure are set in the spool of the pressure compensation valve. A differential pressure generating valve that generates a secondary pressure corresponding to the differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of multiple actuators is provided, and the output pressure is a pressure compensation valve. It may be led to the end of the spool in the opening direction.
[0128]
【The invention's effect】
According to the present invention, in a hydraulic drive device including a turning control system, it is possible to shift to a steady state by accelerating without any jerky feeling of turning operability at the time of starting either turning alone or combining, and turning from turning alone operation. Swivel speed change at the time of transition to composite operation or vice versa is suppressed, and the turn speed can be accelerated without excessively slow compared to other actuators at the time of composite start-up, excellent turning operability and turning independence Can be ensured, and a system that does not cause the problem of an increase in cost and space and a complicated circuit configuration due to the provision of another circuit can be obtained.
[Brief description of the drawings]
FIG. 1 is a circuit diagram showing a hydraulic drive apparatus according to a first embodiment of the present invention.
FIG. 2 is a cross-sectional view showing details of the structure of the pressure compensating valve in the swivel section.
FIG. 3 is a diagram showing a load-dependent characteristic of a pressure compensation valve in a swing section.
FIG. 4 is a diagram showing a lower limit setting function of a target compensation differential pressure by a swing priority spring in a pressure compensation valve of a swing section.
FIG. 5 is a diagram showing an external appearance of a hydraulic excavator in which the hydraulic drive device of the present invention is used.
FIG. 6 is a time chart showing a change in target compensation differential pressure of the pressure compensation valve in the swing section during the swing independent operation.
FIG. 7 is a time chart for explaining the operation of the pressure compensation valve in the swing section when the degree of saturation is large when another actuator is activated during the steady swing rotation.
FIG. 8 is a time chart for explaining the operation of the pressure compensation valve of the swing section when the degree of saturation is small when another actuator is activated during the steady swing rotation.
FIG. 9 is a time chart for explaining the operation of the pressure compensation valve in the turning section when the degree of saturation is large when turning and other actuators are simultaneously activated.
FIG. 10 is a time chart for explaining the operation of the pressure compensation valve in the swing section when the degree of saturation is small when the swing and other actuators are simultaneously activated.
FIG. 11 is a circuit diagram showing a hydraulic drive apparatus according to a second embodiment of the present invention.
FIG. 12 is a view showing an opening area characteristic of a direction switching valve of a turning section.
FIG. 13 is a cross-sectional view showing details of the structure of the pressure compensation valve of the swivel section.
FIG. 14 is a diagram showing the priority characteristic of the flow rate of the swivel section in a saturation state.
FIG. 15 is a circuit diagram showing a hydraulic drive apparatus according to a third embodiment of the present invention.
FIG. 16 is a cross-sectional view showing details of the structure of the pressure compensation valve in the swivel section.
[Explanation of symbols]
1 Hydraulic pump
2-6 Actuator (2: Swing motor)
7-11 Directional switching valve
12-16 Pressure compensation valve
13a-16a Oil chamber (first means)
13b-16b Oil chamber (first means)
18 Pump controller
20-24 Detection line
25-29 signal line
34-36 signal lines
37 signal lines
40 Tilt control actuator
41 Load sensing control valve
50a-50e signal line
51a-51e signal line
52, 52a to 52e Signal line
53, 53a to 52e Signal line
55 Lower limit setting spring (fourth means; biasing means)
60 Setting spring (second means)
200 Undercarriage
201 Revolving body
202 Front work machine
331 Oil chamber (pressure receiving area B1) (second means)
332 Oil chamber (pressure receiving area B3) (third means)
334 Oil chamber (pressure receiving area B1> B3) (third means)
336 Oil chamber (pressure receiving area B2 = B1) (second means)
7A Directional switching valve
12A Pressure compensation valve
55A Turning priority spring (fourth means; biasing means)
57a, 75b Meter-in variable aperture
331A Oil chamber (pressure receiving area B1) (second means)
12B Pressure compensation valve

Claims (7)

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力が前記複数のアクチュエータの最高負荷圧より所定値だけ高くなるようポンプ吐出流量を制御するロードセンシング制御のポンプ制御手段とを備えた油圧駆動装置において、
前記複数の圧力補償弁のうち、前記旋回モータに係わる旋回セクション以外の圧力補償弁に設けられ、前記油圧ポンプの吐出圧力と前記複数のアクチュエータの最高負荷圧との差圧を目標補償差圧として設定する第１手段と、
前記旋回セクションの圧力補償弁に設けられ、その目標補償差圧を設定する第２手段と、
前記複数の圧力補償弁のうち、少なくとも前記旋回セクションの圧力補償弁に設けられ、前記旋回モータの負荷圧が上昇すると、前記第２手段で設定された目標補償差圧を小さくし、旋回セクションの圧力補償弁に負荷依存特性を持たせる第３手段と、
前記旋回セクションの圧力補償弁に設けられ、前記第２手段で設定され、前記第３手段で補正される目標補償差圧の下限が、前記油圧ポンプの吐出圧力と前記複数のアクチュエータの最高負荷圧との差圧が低下したときに前記第１手段により設定される目標補償差圧よりも小さくならないように、前記第２手段で設定され、前記第３手段で補正される目標補償差圧の下限を設定する第４手段とを備えることを特徴とする油圧駆動装置。A plurality of actuators including a hydraulic pump, a swing motor driven by pressure oil discharged from the hydraulic pump, and a plurality of direction switches for controlling the flow rates of pressure oil supplied from the hydraulic pump to the plurality of actuators. A pump, a plurality of pressure compensating valves that respectively control the differential pressure across the plurality of directional control valves, and a pump discharge flow rate so that a discharge pressure of the hydraulic pump is higher than a maximum load pressure of the plurality of actuators by a predetermined value. In a hydraulic drive apparatus including a load sensing control pump control means for controlling,
Among the plurality of pressure compensation valves, provided in a pressure compensation valve other than the swing section related to the swing motor, a differential pressure between a discharge pressure of the hydraulic pump and a maximum load pressure of the plurality of actuators is set as a target compensated differential pressure. A first means for setting;
A second means provided on the pressure compensation valve of the swivel section, for setting the target compensation differential pressure;
Among the plurality of pressure compensation valves, provided at least in the pressure compensation valve of the swing section, when the load pressure of the swing motor increases, the target compensation differential pressure set by the second means is reduced, and A third means for giving the pressure compensation valve a load-dependent characteristic;
The lower limit of the target compensation differential pressure provided in the pressure compensation valve of the swing section, set by the second means, and corrected by the third means is the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators. The lower limit of the target compensation differential pressure that is set by the second means and corrected by the third means so that it does not become smaller than the target compensation differential pressure that is set by the first means when the differential pressure decreases And a fourth means for setting the hydraulic drive device. 請求項１記載の油圧駆動装置において、
前記第２手段は、前記第１手段と同様、前記油圧ポンプの吐出圧力と前記複数のアクチュエータの最高負荷圧との差圧を前記目標補償差圧として設定する手段であり、
前記第４手段は、前記第２手段で設定された目標補償差圧自体の低下と前記第３手段で与えられた負荷依存特性による目標補償差圧の低下の両方に対して下限設定手段として機能することを特徴とする油圧駆動装置。The hydraulic drive device according to claim 1, wherein
The second means, like the first means, is means for setting a differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators as the target compensation differential pressure,
The fourth means functions as a lower limit setting means for both the reduction of the target compensation differential pressure itself set by the second means and the reduction of the target compensation differential pressure due to the load dependence characteristic given by the third means. A hydraulic drive device characterized by that. 請求項１記載の油圧駆動装置において、
前記第２手段は、前記油圧ポンプの吐出圧力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変化しない値を前記目標補償差圧として設定する手段であり、
前記第４手段は、前記第３手段で与えられた負荷依存特性による目標補償差圧の低下に対して下限設定手段として機能することを特徴とする油圧駆動装置。The hydraulic drive device according to claim 1, wherein
The second means is a means for setting, as the target compensation differential pressure, a value that does not change due to a differential pressure between the discharge pressure of the hydraulic pump and the maximum load pressure of the plurality of actuators.
The hydraulic drive apparatus according to claim 4, wherein the fourth means functions as a lower limit setting means with respect to a decrease in the target compensation differential pressure due to the load dependence characteristic given by the third means. 請求項１〜３のいずれか１項記載の油圧駆動装置において、
前記第４手段は、前記第２手段で設定され、前記第３手段で補正される目標補償差圧が所定値に達すると、前記旋回セクションの圧力補償弁のスプールに開け方向の付勢力を付与する付勢手段であることを特徴とする油圧駆動装置。In the hydraulic drive unit according to any one of claims 1 to 3,
The fourth means applies an urging force in the opening direction to the spool of the pressure compensation valve of the swing section when the target compensation differential pressure set by the second means and corrected by the third means reaches a predetermined value. A hydraulic drive device characterized by being an urging means. 請求項４記載の油圧駆動装置において、
前記付勢手段は、前記第２手段で設定され、前記第３手段で補正される目標補償差圧が所定値に達すると、前記旋回セクションの圧力補償弁のスプールに作用し、このスプールを開け方向に付勢する下限設定バネであることを特徴とする油圧駆動装置。The hydraulic drive device according to claim 4, wherein
When the target compensation differential pressure set by the second means and corrected by the third means reaches a predetermined value, the biasing means acts on the spool of the pressure compensation valve of the swing section and opens the spool. A hydraulic drive device comprising a lower limit setting spring biasing in a direction. 請求項１又は２記載の油圧駆動装置において、
前記第４手段は、前記第２手段で設定され、前記第３手段で補正される目標補償差圧に常時補助的な値を付加する付勢手段であり、
前記旋回セクションの方向切換弁は、そのメータイン可変絞りの開口面積が、前記付勢手段で付加される補助的な値の目標補償圧相当分だけ、旋回セクション以外の方向切換弁の開口面積より小さくなるように構成されていることを特徴とする油圧駆動装置。The hydraulic drive device according to claim 1 or 2,
The fourth means is an urging means for always adding an auxiliary value to the target compensation differential pressure set by the second means and corrected by the third means,
The direction switching valve of the turning section has an opening area of the meter-in variable throttle that is smaller than the opening area of the direction switching valve other than the turning section by an amount corresponding to the target compensation pressure of the auxiliary value added by the biasing means. It is comprised so that it may become. The hydraulic drive device characterized by the above-mentioned. 請求項６記載の油圧駆動装置において、
前記付勢手段は、前記旋回セクションの圧力補償弁のスプールの開け方向に常時作用する旋回優先バネであることを特徴とする油圧駆動装置。The hydraulic drive device according to claim 6, wherein
The hydraulic drive device according to claim 1, wherein the biasing means is a turning priority spring that always acts in an opening direction of a spool of the pressure compensating valve of the turning section.
JP34382399A 1998-12-03 1999-12-02 Hydraulic drive Expired - Lifetime JP3853123B2 (en)

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