JP6331010B2 - Hydraulic drive - Google Patents

Hydraulic drive Download PDF

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JP6331010B2
JP6331010B2 JP2014090346A JP2014090346A JP6331010B2 JP 6331010 B2 JP6331010 B2 JP 6331010B2 JP 2014090346 A JP2014090346 A JP 2014090346A JP 2014090346 A JP2014090346 A JP 2014090346A JP 6331010 B2 JP6331010 B2 JP 6331010B2
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pressure
differential pressure
hydraulic pump
sleeve
pump
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JP2015209866A (en
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康治 岡崎
康治 岡崎
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Nachi Fujikoshi Corp
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Description

本発明は、油圧ショベルなどの建設機械及び、各種作業機械に使用される油圧駆動装置に関し、さらに詳細には斜板式可変容量型油圧ポンプ（以下可変ポンプとする）を備えた油圧駆動装置において、特に可変ポンプの吐出圧と複数のアクチュエータの最高負荷圧との実差圧（以下ＰＬＳ）圧という）をある目標差圧（以下Ｐｒ圧）に保つように、可変ポンプの容量を制御するロードセンシング制御の油圧駆動装置に関する。   The present invention relates to a hydraulic drive device used in construction machines such as a hydraulic excavator and various work machines, and more specifically, in a hydraulic drive device including a swash plate type variable displacement hydraulic pump (hereinafter referred to as a variable pump), In particular, load sensing that controls the capacity of the variable pump so that the actual differential pressure (hereinafter referred to as PLS pressure) between the discharge pressure of the variable pump and the maximum load pressure of a plurality of actuators is maintained at a certain target differential pressure (hereinafter referred to as Pr pressure). The present invention relates to a hydraulic drive device for control.

従来、この種のＰｒ圧は可変ポンプと共にエンジンなどの原動機で駆動される固定容量型ポンプ（以下固定ポンプとする）を利用してエンジン回転数の変化に応じて変化するようにした油圧駆動装置において、走行作動時の省エネを図るため、走行作動を検知し、走行作動時のみＰｒ圧が低下する技術がある。
この従来技術では、走行モータ動作時の流量がその他のアクチュエータ（例えば、ブームシリンダーやアームシリンダー）動作時の流量より少ないが、どのアクチュエータを動作させる場合でもＰｒ圧、ＰＬＳ圧が同じため、流量の少ない走行モータ動作時でも流量制御弁での圧力損失が大流量のアクチュエータ動作時と同じ圧力損失となってしまい、エネルギーロスが無駄に大きいという問題に対し、走行操作を検知し、走行動作時のみ、Ｐｒ圧が低く設定されるようにして、走行動作時の圧力損失を低減している。
ただし、先行技術で圧力損失が低減されているのは、エンジン回転数には連動せず、エンジン回転数は中回転から高回転の領域のみである（例えば、特許文献１を参照。）。 Conventionally, this kind of Pr pressure is changed in accordance with changes in engine speed by using a fixed displacement pump (hereinafter referred to as a fixed pump) driven by a prime mover such as an engine together with a variable pump. In order to save energy at the time of traveling operation, there is a technique for detecting the traveling operation and reducing the Pr pressure only during the traveling operation.
In this prior art, the flow rate during operation of the traveling motor is smaller than the flow rate during operation of other actuators (for example, boom cylinders and arm cylinders), but the Pr pressure and PLS pressure are the same regardless of which actuator is operated. Even when a few travel motors are operating, the pressure loss at the flow control valve is the same pressure loss as when a large flow actuator is operating, and the problem is that the energy loss is unnecessarily large. The pressure loss during the traveling operation is reduced by setting the Pr pressure to be low.
However, the pressure loss in the prior art is not linked to the engine speed, and the engine speed is only in the range from medium to high (see, for example, Patent Document 1).

特開２０１１―２４７３０１号公報JP 2011-247301 A

しかしながら、特許文献１では、走行モータ動作時の流量がその他のアクチュエータ（例えば、ブームシリンダーやアームシリンダー）動作時の流量より少ないが、どのアクチュエータを動作させる場合でもＰｒ圧、ＰＬＳ圧が同じため、流量の少ない走行モータ動作時でも流量制御弁での圧力損失が大流量のアクチュエータ動作時と同じ圧力損失となってしまい、エネルギーロスが無駄に大きいという問題に対し、走行操作を検知し、走行動作時のみ、Ｐｒ圧が低く設定されるようにして、走行動作時の圧力損失を低減している。ただし、先行技術で圧力損失が低減されているのは、エンジン回転数には連動せず、エンジン回転数は中回転から高回転の領域のみである。   However, in Patent Document 1, although the flow rate during operation of the traveling motor is smaller than the flow rate during operation of other actuators (for example, boom cylinders and arm cylinders), the Pr pressure and PLS pressure are the same regardless of which actuator is operated. Even when the travel motor with a low flow rate is operating, the pressure loss at the flow control valve is the same pressure loss as when the actuator with a large flow rate is operating, and the travel operation is detected to solve the problem that the energy loss is unnecessarily large. Only when the Pr pressure is set low, the pressure loss during the running operation is reduced. However, in the prior art, the pressure loss is reduced without being linked to the engine speed, and the engine speed is only in the region from medium to high.

従来技術では、図５のようにエンジン回転数が中回転から高回転の場合のみ、走行動作時、Ｐｒ圧はＰｒ´圧のように低下するが低回転ではＰｒ圧のまま変化しない。
そして、一般にエンジン高回転時に実機に必要な走行速度が出るように走行モータの流量調整弁を設定する。従って、エンジン高回転時のＰｒ圧がＰｒ´圧まで低くなった状態で所望の流量となるよう流量調整弁を設定するため開口面積が大きくなる。
しかし、エンジン低回転では、走行動作を検知しても、Ｐｒ圧は低下しないため、高回転時のＰｒ´圧との差が小さく、本来、エンジン低回転時ではアクチュエータの速度がエンジン高回転時に対し一般的には半減するところ、特許文献１では、理論的にそのように大幅に低下することはない。 In the prior art, as shown in FIG. 5, only when the engine speed is from medium to high, Pr pressure decreases like Pr ′ pressure during traveling operation, but Pr pressure remains unchanged at low rotation.
In general, the flow adjustment valve of the travel motor is set so that the travel speed required for the actual machine is obtained at the time of high engine rotation. Accordingly, the opening area is increased because the flow rate adjusting valve is set so that the desired flow rate is obtained in a state where the Pr pressure at the time of high engine rotation is reduced to the Pr ′ pressure.
However, since the Pr pressure does not decrease even when a running operation is detected at a low engine speed, the difference from the Pr ′ pressure at a high engine speed is small. On the other hand, in general, it is halved, but in Patent Document 1, it does not decrease so much theoretically.

よって、実機のオペレータは走行以外のアクチュエータはエンジン回転数によって速度が大幅に変化するが、走行動作についてはエンジン回転数が低回転でも走行速度が低下せず、微操作が困難になってしまう。
さらに、特許文献１における図１０の回路例以外では、走行動作時のＰｒ圧を低下させるため新たに減圧弁や絞りを追加して構成部品を増やすようにしている。
本発明は前記の課題を解決するためになされたもので、エンジン回転数に応じて目標差圧を変更する機能を従来どおり有したまま、走行動作時のみ、目標差圧の絶対値が走行作時以外よりも低くなるようにし、走行動作時の油圧回路内での圧力損失を低減すること、かつ走行動作時でもエンジン回転数に応じてＰｒ圧が変化するようにした差圧減圧弁を有することを特徴とする油圧駆動装置を提供することを目的とする。 Therefore, although the speed of actuators other than traveling varies greatly depending on the engine speed, the actual operator does not decrease the traveling speed even if the engine speed is low, making it difficult to perform fine operations.
Further, except for the circuit example of FIG. 10 in Patent Document 1, in order to reduce the Pr pressure during the running operation, a pressure reducing valve and a throttle are newly added to increase the number of components.
The present invention has been made in order to solve the above-described problems. The absolute value of the target differential pressure is set so that the absolute value of the target differential pressure can be changed only during the traveling operation while maintaining the function of changing the target differential pressure according to the engine speed. It has a differential pressure reducing valve that reduces the pressure loss in the hydraulic circuit during traveling operation and changes the Pr pressure according to the engine speed even during traveling operation. It is an object of the present invention to provide a hydraulic drive device characterized by this.

前記の課題を解決するため、本発明は、
エンジンなどの原動機と、前記エンジンなどの原動機により駆動される可変容量型の油圧ポンプと、前記油圧ポンプから吐出される圧油により駆動される複数のアクチュエータと、前記油圧ポンプから前記複数のアクチュエータに供給される圧油の流量をそれぞれ制御する複数の方向切換弁と、前記複数の方向切換弁の前後差圧をそれぞれ制御する複数の圧力補償弁と、前記油圧ポンプの吐出圧が前記複数のアクチュエータの最高負荷圧よりも目標差圧だけ高くなるロードセンシング制御するポンプ制御手段と、前記油圧ポンプの吐出圧の上限を規制するメインリリーフ弁と、前記複数の圧力補償弁のそれぞれの目標差圧を前記油圧ポンプの吐出圧と前記複数のアクチュエータの最高負荷圧との差圧により設定すると共に、固定ポンプの吐出流量が固定絞りと可変絞りを通過する時の上流圧と下流圧の差圧を検出する差圧減圧弁と、備え
前記差圧減圧弁は、
段付円筒形状のスリーブと、
前記スリーブの一端に嵌着されたガイドと、
前記スリーブの他端に螺着された配管継手と、
前記スリーブに形成した段付孔に嵌着された段付形状のスプールと、
前記配管継手により前記段付孔に嵌着されたスプリングガイドと、
前記スプールの一端に嵌着されたリテーナと、

前記スプリングガイドと前記リテーナとの間に装着されたばね部材と、を備えることを特徴とする。 In order to solve the above problems, the present invention provides:
A prime mover such as an engine, a variable displacement hydraulic pump driven by the prime mover such as the engine, a plurality of actuators 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. A pump control means for load sensing that is higher than the maximum load pressure by a target differential pressure, a main relief valve that regulates an upper limit of the discharge pressure of the hydraulic pump, and a target differential pressure of each of the plurality of pressure compensation valves. and it sets the differential pressure between the maximum load pressure in the discharge pressure and the plurality of actuators of the hydraulic pump, the discharge of the fixed pump A differential pressure reducing valve for detecting the differential pressure of the upstream pressure and the downstream pressure when the amount passes through the fixed throttle and the variable throttle, with the differential pressure reducing valve,
A stepped cylindrical sleeve;
A guide fitted to one end of the sleeve;
A pipe joint is screwed to the other end of the sleeve,
A step-shaped spool fitted into a step hole formed in the sleeve;
A spring guide fitted into the stepped hole by the pipe joint;
A retainer fitted to one end of the spool;

And a spring member mounted between the spring guide and the retainer.

本発明によれば、固定ポンプの吐出路に設置した絞りの前後差圧の変化でエンジン回転数を検出し、走行非動作時にポンプ容量制御の目標差圧Ｐｒ圧を変更し、可変ポンプの吐出圧と最高負荷圧の実差圧も変更する機能は有したままで、走行動作時に目標差圧Ｐｒ´圧を目標走行Ｐｒ圧よりも低い圧力に設定すると共に、Ｐｒ´圧もエンジン回転数に連動して変更する機能も持ち合わせることができる。
さらに、走行動作時の圧力損失は従来技術同様に低減できるうえに、エンジン回転数低回時の走行動作微操作性も確保できるシステムとなる。 According to the present invention, the engine speed is detected by a change in the differential pressure across the throttle installed in the discharge path of the fixed pump, and the target differential pressure Pr pressure of the pump displacement control is changed when the vehicle is not running, so that the variable pump discharge While maintaining the function of changing the actual differential pressure between the pressure and the maximum load pressure, the target differential pressure Pr ′ pressure is set to a pressure lower than the target travel Pr pressure during the travel operation, and the Pr ′ pressure is also set to the engine speed. You can also have a function to change in conjunction.
Further, the pressure loss during the traveling operation can be reduced as in the prior art, and the traveling operation fine operability can be ensured when the engine speed is low.

本発明の第一の実施の形態に係る油圧制御回路を示す。1 shows a hydraulic control circuit according to a first embodiment of the present invention. 図１の油圧制御回路の特性を示す目標ＬＳ差圧―エンジン回転数の関係を示す特性線図である。FIG. 2 is a characteristic diagram showing a relationship between a target LS differential pressure and an engine speed indicating the characteristics of the hydraulic control circuit in FIG. 1. 本発明の第二の実施の形態に係る油圧制御回路である。3 is a hydraulic control circuit according to a second embodiment of the present invention. 図１の差圧減圧弁の概略構造を示す略縦断面図である。FIG. 2 is a schematic longitudinal sectional view showing a schematic structure of a differential pressure reducing valve in FIG. 1. 従来の油圧制御回路の特性を示す目標ＬＳ差圧―エンジン回転数の関係を示す特性線図である。It is a characteristic diagram which shows the relationship of the target LS differential pressure-engine speed which shows the characteristic of the conventional hydraulic control circuit.

以下、本発明に係る油圧駆動装置について、好適な実施の形態を挙げ、添付図面を参照して詳細に説明する。
図１は、本発明の実施の形態に係る油圧駆動装置１０の油圧回路を示し、参照符号１４は油圧駆動装置１０のコントロールバルブを示す。
本発明では、固定ポンプ１３の吐出流量が固定絞り３０ｂと可変絞り３０ａを通過する時の上流圧Ｐｐ１（３１ｃ）と下流圧Ｐｐ２（３１ｂ）の差圧を検出する差圧減圧弁３１で目標差圧Ｐｒ圧（３１ａ）を検出している。
ここで、上流圧Ｐｐ１（３１ｃ）の受圧面積Ａ１、下流圧Ｐｐ２（３１ｂ）の受圧面積をＡ２、目標差圧Ｐｒ圧（３１ａ）の受圧面積をＡ３とする。
差圧減圧弁３１には今回の発明で下流圧３１ｂと同じ方向に作用するバネ３１ｄと走行動作信号圧３１ｆによってバネ３１ｄの荷重（弾発力）Ｆｓｐを変化させる圧力室３１ｅが追加された。走行動作信号圧３１ｆには走行モータ１６、走行モータ１７を動作させるための方向切換弁２６、方向切換弁２７を切り換えるパイロット圧２６ａ、２６ｂ、２７ａ、２７ｂをシャトル弁２８，２９などで高圧選択し、供給する。 DESCRIPTION OF EMBODIMENTS Hereinafter, a hydraulic drive device according to the present invention will be described in detail with reference to the accompanying drawings by giving preferred embodiments.
FIG. 1 shows a hydraulic circuit of a hydraulic drive device 10 according to an embodiment of the present invention, and reference numeral 14 denotes a control valve of the hydraulic drive device 10.
In the present invention, the target pressure difference is detected by the differential pressure reducing valve 31 that detects the differential pressure between the upstream pressure Pp1 (31c) and the downstream pressure Pp2 (31b) when the discharge flow rate of the fixed pump 13 passes through the fixed throttle 30b and the variable throttle 30a. The pressure Pr pressure (31a) is detected.
Here, the pressure receiving area A1 of the upstream pressure Pp1 (31c), the pressure receiving area of the downstream pressure Pp2 (31b) is A2, and the pressure receiving area of the target differential pressure Pr pressure (31a) is A3.
In the present invention, the differential pressure reducing valve 31 is provided with a spring 31d acting in the same direction as the downstream pressure 31b and a pressure chamber 31e for changing the load (elastic force) Fsp of the spring 31d by the traveling operation signal pressure 31f. For the travel operation signal pressure 31f, the pilot pressures 26a, 26b, 27a, 27b for switching the direction switching valve 26 and the direction switching valve 27 for operating the travel motor 16 and the travel motor 17 are selected by the shuttle valves 28, 29, etc. Supply.

よって、走行モータ１６，１７の走行動作（走行モータ作動）時に圧力室３１ｅに走行動作信号圧３１ｆが供給されるとバネ３１ｄが撓んで弾発力がｕｐし、走行非動作時に対し、差圧減圧弁３１で検出されるＰｒ圧３１ａは小さくなる。また、エンジン１１の回転数増加に伴い固定ポンプ１２の吐出流量が増加し、上流圧３１ｃと下流圧３１ｂの差圧が大きくなり、差圧減圧弁３１のスプールバランス位置がより図の上側（バネ３１ｄがたわむ位置）になる。従って、エンジン１１の回転数の上昇に伴い、バネ３１ｄの弾発力が大きくなり、結果として、目標差圧Ｐｒ圧３１ａは回転数が高いほど、上流圧３１ｃと下流圧３１ｂの差圧との圧力差が大きくなる。   Therefore, when the travel operation signal pressure 31f is supplied to the pressure chamber 31e during the travel operation (travel motor operation) of the travel motors 16 and 17, the spring 31d is bent and the elastic force is increased, so that the differential pressure is different from that when the travel is not performed. The Pr pressure 31a detected by the pressure reducing valve 31 decreases. Further, the discharge flow rate of the fixed pump 12 increases with the increase in the number of revolutions of the engine 11, the differential pressure between the upstream pressure 31c and the downstream pressure 31b increases, and the spool balance position of the differential pressure reducing valve 31 is more upward in the drawing (spring 31d bends). Therefore, as the rotational speed of the engine 11 increases, the spring force of the spring 31d increases. As a result, the higher the rotational speed of the target differential pressure Pr pressure 31a, the higher the differential pressure between the upstream pressure 31c and the downstream pressure 31b. The pressure difference increases.

走行動作時の特性は、図２に示すようになり、従来技術のようにエンジン回転数が中回転から高回転のみ目標差圧Ｐｒ圧３１ａが低下するのではなく、エンジン低回転から高回転まで全域で目標差圧Ｐｒ圧３１ａが低下する特性となる。
可変ポンプ１２の吐出量を調整するＬＳ弁２３には目標差圧Ｐｒ圧（３１ａ）とコントロールバルブ１０の内部のポンプ圧１２ｂと高負荷圧２１ａを差圧減圧弁３１に導き、実差圧ＰＬＳ２２を比較し、実差圧ＰＬＳ２２は目標差圧Ｐｒ圧（３１ａ）と等しくなるように可変ポンプ１２の吐出量が調整される。 The characteristics during the running operation are as shown in FIG. 2, and the target differential pressure Pr pressure 31a does not decrease only when the engine speed is medium to high, as in the prior art, but from low engine speed to high engine speed. The target differential pressure Pr pressure 31a decreases over the entire area.
The target differential pressure Pr pressure (31a) , the pump pressure 12b inside the control valve 10 and the high load pressure 21a are guided to the differential pressure reducing valve 31 to the LS valve 23 for adjusting the discharge amount of the variable pump 12 , and the actual differential pressure PLS22 is obtained. And the discharge amount of the variable pump 12 is adjusted so that the actual differential pressure PLS22 becomes equal to the target differential pressure Pr pressure (31a) .

極端な例として考えると、走行非動作時バネ３１ｄはたわみがない状態とし、走行動作時のみバネ３１ｄがたわみ，バネ力が差圧減圧弁３１に作用する差圧減圧弁３１の力の釣り合い式で表すと、
Ｐｐ１×Ａ１＝Ｐｐ２×Ａ＋Ｐｒ×Ａ３＋Ｆｓｐ（バネ３１ｄの弾発力）
整理すると、
Ｐｒ×Ａ３＝Ｐｐ２×Ａ１−Ｐｐ２×Ａ２−Ｆｓｐ
各圧力の受圧面積Ａ１、Ａ２、Ａ３の面積の比は目的に応じ設定するが、差圧減圧弁として最も一般的な場合で考え、Ａ１＝Ａ２＝Ａ３と設定すると、
Ｐｒ×Ａ１＝Ｐｐ１×Ａ１−Ｐｐ２×Ａ１−Ｆｓｐ
∴Ｐｒ＝Ｐｐ１−Ｐｐ２−Ｆｓｐ／Ａ１となる。

走行非動作時：目標差圧Ｐｒ=Ｐｐ１−Ｐｐ２（Ｆｓｐ／Ａ１＝０なので） ・・・（１）

走行動作時： 目標差圧Ｐｒ´=Ｐｐ１−Ｐｐ２−Ｆｓｐ／Ａ１ ・・・（２）
となる。
よって、走行動作時 実差圧ＰＬＳ２２＝目標差圧Ｐｒ´であり、走行時の圧力損失が低減できる結果となる。しかもエンジン回転数低回転から高回転まで全領域で走行非動作時の目標差圧Ｐｒ圧同様Ｐｒ´圧も変化することが可能となる。 Considering as an extreme example, the spring 31d is in a state in which there is no deflection when the vehicle is not operating, and the spring 31d is bent only during the traveling operation, and the balance of the force of the differential pressure reducing valve 31 where the spring force acts on the differential pressure reducing valve 31. In terms of
Pp1 × A1 = Pp2 × A + Pr × A3 + Fsp (elasticity of spring 31d)
Organize
Pr * A3 = Pp2 * A1-Pp2 * A2-Fsp
The ratio of the pressure receiving areas A1, A2 and A3 of each pressure is set according to the purpose, but in the most general case as a differential pressure reducing valve, if A1 = A2 = A3 is set,
Pr * A1 = Pp1 * A1-Pp2 * A1-Fsp
∴Pr = Pp1-Pp2-Fsp / A1.

When not running: Target differential pressure Pr = Pp1-Pp2 (since Fsp / A1 = 0) (1)

During travel operation: Target differential pressure Pr '= Pp1-Pp2-Fsp / A1 (2)
It becomes.
Therefore, the actual differential pressure PLS22 at the time of running operation = the target differential pressure Pr ′, and the pressure loss during running can be reduced. In addition, the Pr ′ pressure can be changed in the same manner as the target differential pressure Pr pressure when the vehicle is not running in the entire region from the low engine speed to the high engine speed.

図４は他のコントロールバルブ１５の回路例を示し、図４中、図１のコントロールバルブ１４の構成要件と同一の構成要件は同一符号を付して詳細な説明を省略する。
よって、図１のコントロールバルブ１４の代わりに図４のコントロールバルブ１５に置き換えても同じ機能が得られる。
図４について説明する。図１のコントロールバルブ１０との相違点は、固定ポンプ１３の吐出ライン下流のパイロット圧３２が入力されたラインを分岐し、絞り３３を配置する点にある。
絞り３３の下流は走行モータ１６，１７の操作用方向切換弁２６，２７に直結した信号ライン遮断回路５１，５２に直列に接続され、コントロールバルブ１５のタンク内でタンクポート（Ｔライン）に接続されていることである。 FIG. 4 shows another circuit example of the control valve 15. In FIG. 4, the same constituent elements as those of the control valve 14 of FIG.
Therefore, the same function can be obtained by replacing the control valve 14 of FIG. 1 with the control valve 15 of FIG.
FIG. 4 will be described. A difference from the control valve 10 in FIG. 1 is that a line to which a pilot pressure 32 downstream of the discharge line of the fixed pump 13 is inputted is branched and a throttle 33 is arranged.
The downstream of the throttle 33 is connected in series to signal line cutoff circuits 51 and 52 directly connected to the operation direction switching valves 26 and 27 of the travel motors 16 and 17, and is connected to the tank port (T line) in the tank of the control valve 15. It has been done.

また、絞り３３の下流は分岐され、走行動作信号圧３１ｆに供給される。走行非動作時は絞り３３の下流はタンクへ連通しており、０．５ＭＰａ以下の低圧になっている。
そのため、走行動作信号圧３１ｆは０．５ＭＰａ以下でバネ３１ｄは弾発力が弱い状態である。走行動作時に方向切換弁２６あるいは２７が切り換わる、または方向切換弁２６、２７が同時に切り換わり、信号ライン遮断回路５１，あるいは５２が遮断するポジションに切りかわると、パイロット圧３２のリリーフ圧まで絞り３３の下流圧が昇圧し、走行動作信号圧３１ｆに供給され、バネ３１ｄが撓み、弾発力がｕｐし、その結果、目標差圧Ｐｒ圧３１ａが低下することになる。 Further, the downstream side of the throttle 33 is branched and supplied to the travel operation signal pressure 31f. When the vehicle is not running, the downstream side of the throttle 33 communicates with the tank and has a low pressure of 0.5 MPa or less.
Therefore, the traveling operation signal pressure 31f is 0.5 MPa or less and the spring 31d is in a state where the elastic force is weak. When the direction switching valve 26 or 27 is switched during traveling operation or when the direction switching valves 26 and 27 are switched at the same time and the signal line shut-off circuit 51 or 52 is switched to the shut-off position, the pilot pressure 32 is reduced to the relief pressure. The downstream pressure of 33 is increased and supplied to the travel operation signal pressure 31f, the spring 31d is bent, the elasticity is increased, and as a result, the target differential pressure Pr pressure 31a is decreased.

本発明の実施に形態に係る油圧駆動装置１０の差圧減圧弁３１の概略構造を図３に示す。
差圧減圧弁３１は段付円筒形状のスリーブ３２２と、前記スリーブ３２２の一端（図３で左端）に嵌着されたガイド３２１と、前記スリーブ３２２の他端（図３で右端）に螺着された配管継手３１５と、前記スリーブ３２２に形成した段付孔３２３に嵌着された段付形状のスプール３１２と、前記配管継手３１５により前記段付孔３２３に嵌着されたスプリングガイド３１４と、前記スプール３１２の一端（図３で右端）に嵌着されたリテーナ３１３と、前記スプリングガイド３１４の突起部３１４ａと前記リテーナ３１３との間に装着されたばね部材３１ｄと、を備える。 FIG. 3 shows a schematic structure of the differential pressure reducing valve 31 of the hydraulic drive device 10 according to the embodiment of the present invention.
The differential pressure reducing valve 31 includes a stepped cylindrical sleeve 322, a guide 321 fitted to one end (the left end in FIG. 3) of the sleeve 322, and the other end (the right end in FIG. 3) of the sleeve 322. A pipe joint 315, a stepped spool 312 fitted in a stepped hole 323 formed in the sleeve 322, a spring guide 314 fitted in the stepped hole 323 by the pipe joint 315, A retainer 313 fitted to one end of the spool 312 (the right end in FIG. 3), and a spring member 31 d fitted between the protrusion 314 a of the spring guide 314 and the retainer 313.

一方、スリーブ３２２には、ピストン３１１に対応位置に直径方向に横穴４７が穿設されており、さらに、目標差圧Ｐｒに連通する横穴４９、４４、４３が直径方向に穿設されている。
なお、ばね部材３１ｄはスプリングガイド３１４の突起部３１４ａとスリーブ３１２の他端（図３で右端）の細部３１２ａに嵌着されたリテーナ３１３との間に装着されている。
参照符号３１６乃至３１９はスリーブ３２２の外周に嵌着されたシール部材、例えばＯリングを示しており、差圧減圧弁３１を弁本体（図示しない）に装着した際の該差圧減圧弁３１からの油漏れを防止している。 On the other hand, the sleeve 322 is provided with a lateral hole 47 in the diameter direction at a position corresponding to the piston 311 , and further, lateral holes 49, 44, 43 communicating with the target differential pressure Pr are formed in the diameter direction.
The spring member 31d is mounted between the protrusion 314a of the spring guide 314 and the retainer 313 fitted to the detail 312a at the other end (right end in FIG. 3) of the sleeve 312.
Reference numerals 316 to 319 denote seal members fitted to the outer periphery of the sleeve 322, for example, an O-ring. From the differential pressure reducing valve 31 when the differential pressure reducing valve 31 is attached to a valve body (not shown). Prevents oil leakage.

次に差圧減圧弁３１の作動について説明する。
ピストン３１０の左端面４１に、上流圧Ｐｐ１（３１ｃ）が加圧されると、図３の矢印Ｘ方向への力がピストン３１０の右端面と接してスプール３１２に作用する。また、下流圧Ｐｐ２（３１ｂ）がスリーブ３２２の横穴４３を経由して、スプール３１２の外径大部と外径小部の段差とスリーブ３２２の内径部で形成される圧力室４２に加圧され、図３で矢印Ｙ方向への力がスプール３１２に作用する。 Next, the operation of the differential pressure reducing valve 31 will be described.
When the upstream pressure Pp1 (31c) is applied to the left end surface 41 of the piston 310 , a force in the direction of arrow X in FIG. 3 contacts the right end surface of the piston 310 and acts on the spool 312. Further, the downstream pressure Pp2 (31b) is pressurized through the lateral hole 43 of the sleeve 322 to the pressure chamber 42 formed by the step between the outer diameter large portion and the outer diameter small portion of the spool 312 and the inner diameter portion of the sleeve 322. In FIG. 3, a force in the direction of arrow Y acts on the spool 312.

また、バネ３１ｄの弾発力が同じく図３の矢印Ｙ方向へスプール３１２に作用する。これらの入力でスプール３１２は左右方向に動き、例えばスプール３１２が右へ動くと、スプール３１２の右端面の圧力室４６はスプール３１２の内部穴４５を通してスリーブ３２２の横穴４４を経由して下流圧Ｐｐ２（３１ｂ）と連通する。
逆にスプール３１２が左に動くと、スプール３１２の右端面の圧力室４６はスプール３１２の内部穴４５を通して、スリーブ３２２の横穴４７を経由して、ＤＲポート３１ｇに連通して、スプール３１２に作用する力が釣り合うような目標差圧Ｐｒ（３１ａ）を出力する。 Further, the elastic force of the spring 31d similarly acts on the spool 312 in the direction of arrow Y in FIG. With these inputs, the spool 312 moves in the left-right direction. For example, when the spool 312 moves to the right, the pressure chamber 46 on the right end surface of the spool 312 passes through the internal hole 45 of the spool 312 and the downstream pressure Pp2 via the lateral hole 44 of the sleeve 322. It communicates with (31b) .
Conversely, when the spool 312 moves to the left, the pressure chamber 46 on the right end surface of the spool 312 communicates with the DR port 31g through the inner hole 45 of the spool 312 and the lateral hole 47 of the sleeve 322 to act on the spool 312. The target differential pressure Pr (31a) is output so that the force to be balanced is balanced .

配管継手３１５に走行動作信号圧３１ｆが加圧されると、ピストン３１１が左方向に押され、スリーブ３２２の内部段差４８に押しつけられる。これにより、ピストンＢ３１４が移動した距離分バネ３１ｄの弾発力が高くなる。
よって、ピストン３１０の左端面４１の受圧面積と受圧面積室４２のスプール３１２の外径大部と外径小部の段差部受圧面積、及び受圧室４６のスプール３１２の外径小部の受 When the travel operation signal pressure 31 f is applied to the pipe joint 315, the piston 311 is pushed leftward and is pushed against the internal step 48 of the sleeve 322. Thereby, the elasticity of the spring 31d is increased by the distance that the piston B 314 has moved.
Accordingly, the pressure receiving area of the left end surface 41 of the piston 310, the stepped portion pressure receiving area of the outer diameter large portion and the outer diameter small portion of the spool 312 of the pressure receiving area chamber 42, and the receiving of the small outer diameter portion of the spool 312 of the pressure receiving chamber 46 are received.

１０ 油圧駆動装置 １１ エンジン
１２ 可変ポンプ １３ 固定ポンプ
１４、１５ コントロールバルブ １６、１７ 走行モータ
２６，２７ 方向切換弁 ２８，２９ シャトル弁
３０ａ 可変絞り ３０ｂ 固定絞り
３１ａ 目標差圧Ｐｒ ３１ｂ 下流圧Ｐｐ２
３１ｃ 上流圧Ｐｐ１ ３１ｄ ばね
３１ｅ 圧力室 ３１ｆ 走行動作信号圧
３１ 差圧減圧弁 ３２ パイロット圧
３３ 絞り ５１，５２ 遮断回路
３１０，３１１ ピストン ３１２ スプール
３１３ リテーナ ３１４ スプリングガド
３１５ 配管継手 ３２１ ガイド
３２２ スリーブ ３２３ 段付孔 DESCRIPTION OF SYMBOLS 10 Hydraulic drive device 11 Engine 12 Variable pump 13 Fixed pump 14, 15 Control valve 16, 17 Traveling motor 26, 27 Direction switching valve 28, 29 Shuttle valve 30a Variable throttle 30b Fixed throttle 31a Target differential pressure Pr 31b Downstream pressure Pp2
31c Upstream pressure Pp1 31d Spring 31e Pressure chamber 31f Running operation signal pressure 31 Differential pressure reducing valve 32 Pilot pressure 33 Throttle 51, 52 Shut-off circuit 310, 311 Piston 312 Spool 313 Retainer 314 Spring gad 315 Pipe joint 321 Guide 322 Sleeve 323 Stepped Hole

Claims (1)

エンジンなどの原動機と、前記エンジンなどの原動機により駆動される可変容量型の油圧ポンプと、前記油圧ポンプから吐出される圧油により駆動される複数のアクチュエータと、前記油圧ポンプから前記複数のアクチュエータに供給される圧油の流量をそれぞれ制御する複数の方向切換弁と、前記複数の方向切換弁の前後差圧をそれぞれ制御する複数の圧力補償弁と、前記油圧ポンプの吐出圧が前記複数のアクチュエータの最高負荷圧よりも目標差圧だけ高くなるロードセンシング制御するポンプ制御手段と、前記油圧ポンプの吐出圧の上限を規制するメインリリーフ弁と、前記複数の圧力補償弁のそれぞれの目標差圧を前記油圧ポンプの吐出圧と前記複数のアクチュエータの最高負荷圧との差圧により設定すると共に、固定ポンプの吐出流量が固定絞りと可変絞りを通過する時の上流圧と下流圧の差圧を検出する差圧減圧弁と、備え
前記差圧減圧弁は、
段付円筒形状のスリーブと、
前記スリーブの一端に嵌着されたガイドと、
前記スリーブの他端に螺着された配管継手と、
前記スリーブに形成した段付孔に嵌着された段付形状のスプールと、
前記配管継手により前記段付孔に嵌着されたスプリングガイドと、
前記スプールの一端に嵌着されたリテーナと、
前記スプリングガイドと前記リテーナとの間に装着されたばね部材と、
を、備えることを特徴とする油圧駆動装置。 A prime mover such as an engine, a variable displacement hydraulic pump driven by the prime mover such as the engine, a plurality of actuators 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. A pump control means for load sensing that is higher than the maximum load pressure by a target differential pressure, a main relief valve that regulates an upper limit of the discharge pressure of the hydraulic pump, and a target differential pressure of each of the plurality of pressure compensation valves. and it sets the differential pressure between the maximum load pressure in the discharge pressure and the plurality of actuators of the hydraulic pump, the discharge of the fixed pump A differential pressure reducing valve for detecting the differential pressure of the upstream pressure and the downstream pressure when the amount passes through the fixed throttle and the variable throttle, with the differential pressure reducing valve,
A stepped cylindrical sleeve;
A guide fitted to one end of the sleeve;
A pipe joint is screwed to the other end of the sleeve,
A step-shaped spool fitted into a step hole formed in the sleeve;
A spring guide fitted into the stepped hole by the pipe joint;
A retainer fitted to one end of the spool;
A spring member mounted between the spring guide and the retainer;
A hydraulic drive device comprising:
JP2014090346A 2014-04-24 2014-04-24 Hydraulic drive Active JP6331010B2 (en)

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