WO2017067177A1 - 一种双联轴向柱塞泵恒功率调节系统及其应用 - Google Patents

一种双联轴向柱塞泵恒功率调节系统及其应用 Download PDF

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Publication number
WO2017067177A1
WO2017067177A1 PCT/CN2016/084716 CN2016084716W WO2017067177A1 WO 2017067177 A1 WO2017067177 A1 WO 2017067177A1 CN 2016084716 W CN2016084716 W CN 2016084716W WO 2017067177 A1 WO2017067177 A1 WO 2017067177A1
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WIPO (PCT)
Prior art keywords
spring
converter
inner spring
plunger pump
outer spring
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PCT/CN2016/084716
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English (en)
French (fr)
Inventor
万丽荣
逯振国
曾庆良
张鑫
钟佩思
王亮
王成龙
江守波
孟昭胜
李伟民
许德山
王认辉
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山东科技大学
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Application filed by 山东科技大学 filed Critical 山东科技大学
Priority to AU2016343379A priority Critical patent/AU2016343379B2/en
Priority to CA2974845A priority patent/CA2974845C/en
Publication of WO2017067177A1 publication Critical patent/WO2017067177A1/zh

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
    • F04B49/22—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves

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  • the invention relates to a double axial piston pump constant power regulating system and an application thereof, and belongs to the technical field of hydraulic transmission components.
  • the plunger pump is mainly used in high pressure, high flow and high power working conditions
  • Double axial piston pumps are usually used with prime movers (such as diesel engines). Brand models commonly used in the market, such as K3V series piston pumps designed and manufactured by Kawasaki Heavy Industries Co., Ltd., use constant power regulation systems under working conditions. Ensure that the plunger pump makes full use of the output power of the prime mover, and avoids the prime mover in an overloaded working state, achieving the purpose of saving energy and improving work efficiency; the working principle of the constant power regulating system is to collect the working load pressure signal, and preset The parallel spring balances the force, and the displacement signal of the output is amplified to adjust the output displacement of the plunger pump. Since the two preset parallel springs have different free lengths and different mounting positions, the working time is divided into The three stages of uncompressed, single compression, and parallel compression.
  • the present invention provides a dual axial piston pump constant power regulation system that exhibits more spring rate coefficients without increasing the number of springs.
  • the present invention also provides a regulator comprising the above dual axial piston pump constant power regulation system.
  • the present invention also provides a dual axial piston pump comprising the above regulator.
  • a dual-axis axial piston pump constant power regulating system comprises a main valve body, an outer spring adjusting screw, an inner spring adjusting screw, an outer spring adjusting nut and an inner spring adjusting nut
  • the main valve body comprises an outer spring, an inner spring, and a compensation
  • the rod, the compensating spool and the compensating piston are provided with a hydraulic side cover and a spring side cover at both ends of the main valve body
  • the outer spring adjusting screw is mounted on the main valve body through the spring side cover and fixed by the outer spring adjusting nut
  • the spring adjusting screw is fixed through the outer spring adjusting screw and is fixed by the inner spring adjusting nut
  • the main valve body further comprises an inner spring converter, an outer spring converter and a spring seat, and the inner spring is located in the inner spring converter and the spring seat
  • the inner spring converter penetrates the outer spring and the outer spring converter, one end of the outer spring is in contact with the outer spring adjusting screw, the other end is in contact with the outer spring converter, and the outer spring converter is
  • the inner spring converter has a cylindrical shape
  • the cylinder includes a cavity
  • the inner spring is disposed in the cavity, and one end of the inner spring adjusting screw passes through the inner spring converter and the inner spring.
  • the outer spring converter is a ring having a stepped hole
  • the inner spring converter has an outwardly extending flange at one end edge thereof, and the outer spring converter is in contact with the inner spring converter.
  • the flange is located in the stepped bore.
  • one side of the spring seat is provided with a tapered groove
  • the other side is provided with a boss
  • a groove for accommodating the flange is arranged on the outer side of the boss, and the boss is embedded in the inner spring, the compensation One end of the rod is located in the tapered recess.
  • a regulator includes the above-described constant power regulation system, a negative flow feedback adjustment system, and a valve control cylinder position following system
  • the negative flow feedback adjustment system includes a negative flow feedback cylinder
  • the valve control cylinder position following system includes The servo valve, the feedback lever, the servo plunger and the differential cylinder; the compensation rod and the negative flow feedback cylinder are respectively connected with the servo valve for controlling the movement of the spool of the servo valve, and the servo plunger is disposed in the differential cylinder and passes through the feedback rod Connected to the servo valve, the servo plunger controls the spool movement of the servo valve through the feedback lever.
  • a dual axial piston pump comprising the regulator described above.
  • a working method of a double axial piston pump comprising the following steps,
  • the outlet pressure of the plunger pump also increases, and the force on the compensating spool is gradually increased.
  • the compensating spool generates a leftward force to overcome the preloading force of the inner spring and the outer spring to the right, the compensating rod shifts to the left and correspondingly pushes the inner spring converter to move to the left, and the compensating lever moves to the left to drive the servo valve.
  • the large cavity end of the servo plunger is connected to the high pressure oil, and the servo plunger is displaced to the right, thereby reducing the inclination of the swash plate of the plunger pump and reducing the displacement of the plunger pump;
  • the compensating lever continues to move to the left until the inner spring converter contacts the inner spring adjusting screw;
  • the compensating lever continues to move to the left.
  • the spring seat and the outer spring converter come into contact, the flange and the stepped hole are gradually disengaged from contact, and the outer spring converter is separately moved to the left.
  • the spring seat simultaneously compresses the inner spring and the outer spring;
  • the invention improves the constant power adjustment system inside the double axial piston pump, and designs a brand new spring assembly mechanism-internal-parallel converter (internal spring converter and outer spring converter) at the inner spring and the outer spring.
  • the series-parallel converter can make the two springs exhibit four kinds of spring stiffness characteristics during the compression process, so that the four comprehensive stiffness coefficients exhibited by the two springs in the constant power adjustment process, the characteristic curve and the theoretical constant power hyperbolic curve
  • the degree of fitting is higher; the mechanism can greatly improve the energy utilization rate of the plunger pump pressure load in the range of 15 to 25 MPa; effectively avoiding the plunger pump and the prime mover working in an overload state and improving the service life of the hardware.
  • Figure 1 is a pressure load-output flow characteristic curve
  • FIG. 2 is a schematic diagram of the regulator
  • 3 is a schematic structural view of a constant power regulation system
  • Figure 4 is a schematic structural view of a compensation rod
  • Figure 5 is a schematic structural view of a compensating spool
  • Figure 6a is a front view of the spring seat
  • Figure 6b is a right side sectional view of the spring seat
  • Figure 7a is a front view of the inner spring converter
  • Figure 7b is a left side cross-sectional view of the inner spring converter
  • Figure 8a is a front view of the outer spring converter
  • Figure 8b is a left side cross-sectional view of the outer spring converter
  • Figure 9 is a state diagram when the inner spring and the outer spring are connected in series and compressed
  • Figure 10 is a state diagram when the inner spring is compressed
  • Figure 11 is a view showing a state in which the inner spring is compressed and the inner spring and the outer spring are connected in parallel to start compression;
  • Figure 12 is a state diagram in which the inner spring and the outer spring are connected in parallel at the end of compression
  • Figure 13 is a functional block diagram of a double axial piston pump of the present invention.
  • the embodiment provides a dual-axis axial piston pump constant power regulation system, and the constant power regulation system is a constant power regulation system in the regulator on the plunger pump, and the specific structure thereof is shown in FIG. 3, in the main valve body.
  • the 001 is mainly composed of a spring force compensating mechanism 7 and a hydraulic pressure compensating mechanism 8, wherein the spring force compensating mechanism 7 is composed of an inner spring adjusting screw 701, an inner spring adjusting nut 702, an outer spring adjusting screw 703, an outer spring adjusting nut 704, and a spring. Side cover 705, outer spring 706, inner spring converter 707, inner spring 708, outer spring converter 709, spring seat 710, sealing ring, etc. Piece of composition.
  • the hydraulic pressure compensating mechanism 8 is composed of a compensating rod 801, a compensating valve core 802, a compensating valve sleeve 803, a compensating plug 804, a compensating piston 805, and a seal ring.
  • a spring side cover 705 and a hydraulic side cover 002 are disposed at both ends of the main valve body 001.
  • the inner spring 708 is located between the inner spring converter 707 and the spring seat 710, and the inner spring converter 707 penetrates the outer spring 706 and the outer spring converter 709.
  • One end of the outer spring 706 is in contact with the spring side cover 705, the other end is in contact with the outer spring converter 709, the outer spring converter 709 is in mating contact with the inner spring converter 707, the inner spring converter 707 and the inner spring adjusting screw 701, There is a gap between the spring seats 710.
  • the outer spring adjusting screw 703 is mounted on the main valve body 001 by the spring side cover 705 and fixed by the external spring adjusting nut 704 screw connection;
  • the inner spring adjusting screw 701 is in the shape of a stepped shaft, which includes a shoulder, the inner spring adjusting screw 701 One end passes through the outer spring adjusting screw 703, the inner spring converter 707 and the inner spring 708, and the other end is fixed by the inner spring adjusting nut 702 screwing, the shoulder of the inner spring adjusting screw 701 and the inner spring converter 707 and the outer spring are adjusted A gap is left between the screws 703, and the preload of the outer spring and the inner spring is adjusted by adjusting the amount of entry of the outer spring screw and the inner spring screw in the main valve body.
  • the inner spring converter has a cylindrical shape. As shown in FIG. 7a and FIG. 7b, the cylinder includes a cavity and a mounting hole connecting the cavity, and the inner spring is disposed on the left inner inner left contact seat 7073. One end of the inner spring adjusting screw 701 passes through the fitting hole 7072 of the inner spring converter and is placed in the inner spring 708.
  • the outer spring converter 709 is a ring having a stepped hole, as shown in Figs. 8a and 8b, an outwardly extending flange 7074 is provided at one end edge of the cylinder, and the outer spring converter 709 and the inner spring converter are provided.
  • the flange 7074 is located in the stepped hole 7093.
  • one side of the spring seat 710 is provided with a tapered recess 7104, the other side is provided with a boss 7101, and a groove 7105 for accommodating the flange 7074 is disposed outside the boss 7101.
  • the boss 7101 is embedded in the inner spring 708.
  • the structure of the compensating rod 801 is as shown in FIG. 4, and includes a tapered arc-shaped plug 8011, an annular groove 8012 and a compensating rod tailstock 8013.
  • the tapered arc-shaped head 8011 at the left end of the compensating rod is located in the tapered recess 7104.
  • annular groove 8012 is connected to the spool of the servo valve 10 in the valve-controlled cylinder position following system, and the compensating rod tailstock 8013 at the right end is subjected to the force of the compensating spool arc-shaped head 8021 at the left end of the compensating spool 802.
  • the spring seat 710 is subjected to two spring forces at the left end, wherein the boss 7101 is engaged with the inner circle of the inner spring 708, and the inner spring right contact seat 7102 is in contact with the right end surface of the inner spring.
  • the parallel pressing surface 7103 is in mating contact with the parallel compression surface 7092 of the outer spring converter 709, and the tapered recess 7104 is in mating contact with the tapered circular arc head 8011 of the compensating rod;
  • the inner spring 708, the outer spring 706, the inner spring converter 707 and the outer spring converter 709 together form a series-parallel converter.
  • the distance between the limiting end surface 7071 of the inner spring converter 707 and the shoulder of the inner spring adjusting screw 701 determines the pressure range in which the inner spring 708 and the outer spring 706 are in the series working state, and the fitting hole 7072 is fitted with the inner spring adjusting screw 701.
  • the spring left contact seat 7073 is in contact with the left end surface of the inner spring, and the flange 7074 is in mating contact with the stepped hole 7093;
  • the outer spring right contact seat 7091 of the outer spring converter 709 is in contact with the right end surface of the outer spring 706, the parallel compression surface 7092 and the parallel pressing surface 7103 can be in mating contact, and the stepped hole 7093 is fitted with the flange 7074;
  • the compensating spool 802 is a stepped annular structure located in the compensating valve sleeve 803. The structure is as shown in FIG. 5.
  • the compensating spool arc head 8021 at the left end is in contact with the compensating rod tailstock 8013, and the two annular pumps in the middle are arranged.
  • the active surface 8022 and the other pumping surface 8023 are respectively provided with (double axial piston pump) outlet pressures P 1 and P 2 of the front pump 3 and the rear pump 2 , and the right-end compensating spool tailstock 8024 is subjected to the compensation piston 805.
  • the force compensating piston 805 is placed in the compensating valve sleeve 803 through the outer casing compensation plug 804, and the compensating valve sleeve 803 is located in the main valve body 001.
  • the left end of the compensation piston 805 is the low pressure oil connected to the oil tank, and the hydraulic pressure on the right end is from the electromagnetic proportional pressure reducing valve outlet pressure P f of the electromagnetic proportional pressure reducing valve 9 on the plunger pump.
  • the constant power regulation system designed a novel spring assembly mechanism-series-parallel converter (internal spring converter and outer spring converter) based on the traditional constant power regulation system, and the series-parallel conversion
  • the two springs can exhibit four spring stiffness characteristics during the compression process, so that the four integrated stiffness coefficients exhibited by the two springs have a higher degree of fit to the theoretical constant power hyperbola during constant power regulation.
  • the mechanism can greatly improve the energy utilization rate of the plunger pump pressure load in the range of 15 to 25 MPa; effectively avoiding the plunger pump and the prime mover working in an overload state and improving the service life of the hardware.
  • the embodiment provides a regulator for use on a double axial piston pump, and the regulator includes the dual axial piston pump constant power adjustment system described in Embodiment 1 (by spring force compensation mechanism 7, hydraulic pressure
  • the compensation mechanism 8 is composed of a negative flow feedback adjustment system (consisting of the negative flow feedback cylinder 6) and a set of valve control cylinder position following system (mainly by the differential cylinder 5, the servo valve 10, the feedback rod 11, the servo plunger 12 composition).
  • the negative flow feedback cylinder 6 is connected with the servo valve 10 in the valve control cylinder position following system for controlling the spool movement of the servo valve 10, and the constant power adjustment system passes the compensation rod 801 and the valve control cylinder of the hydraulic compensation mechanism 8 thereof.
  • the servo valve 10 in the position following system is connected, and the movement of the compensating rod is used to control the movement of the spool of the servo valve 10.
  • the servo plunger 12 is disposed in the differential cylinder 5 and connected to the servo valve 10 through the feedback lever 11, and the servo The plunger 12 controls the spool movement of the servo valve 10 through the feedback rod 11, and there is no connection between the constant power regulation system and the negative flow feedback system, and the constant power regulation system and the negative flow feedback system are connected. Other control methods ensure that only one is working.
  • the embodiment provides a double axial piston pump, the structure is as shown in FIG. 2, which includes a pilot pump 1, a main pump (front pump 3 and rear pump 2), an electromagnetic proportional pressure reducing valve 9, and a front pump 3
  • a regulator as described in Example 2 is mounted on each of the rear pump 2.
  • the double axial piston pump is usually connected with the prime mover.
  • the prime mover is selected as a diesel engine, and the prime mover 4 drags two series-connected plunger through-shaft rear pump 2 through the input shaft.
  • the front pump 3 and the pilot pump 1 installed at the rear, wherein the plunger pump supplies high-pressure power oil to the working actuator, and the pilot gear pump supplies high-pressure oil to the control oil passage.
  • the front pump 3 and the rear pump 2 are respectively configured with the regulator 1 and the regulator 2 as described in the second embodiment.
  • the electromagnetic proportional pressure reducing valve 9 can obtain different electromagnetic proportional pressure reducing valve outlet pressures P f according to the change of the input current i, thereby setting the magnitude of the plunger pump working condition power W C .
  • the piston pump operating condition power W C should be smaller than the prime mover rated power W M .
  • the plunger pump regulator will pass the constant power regulation system. The function is to reduce the output flow of the plunger pump, thereby avoiding the overload operation of the prime mover 4, and realizing the self-protection mode of "heavy heavy load".
  • the constant power regulation system uses the load pressures P 1 and P 2 of the front pump 3 and the rear pump 2 as control signals to adjust the output displacement of the plunger pump.
  • This control method is called a cross power control strategy, and the advantage is that the entire operation can be guaranteed.
  • the total power output of the hydraulic pump is constant, and when the power required by one of the pumps is reduced, the other pump can automatically utilize the remaining power to make full use of the prime mover 4 power.
  • the purpose of constant power control is to ensure that the plunger pump can establish a good matching relationship with the output power of the prime mover 4 during the working process by adjusting the displacement of the plunger pump, making full use of the power energy and ensuring The output power of prime mover 4 does not exceed the rated power, extending the service life of the electro-hydraulic system.
  • the compensating rod 801 The displacement x R is generated to the left, and the valve-controlled cylinder position following system acts through the servo valve 10, so that the large cavity end of the servo plunger 12 is connected to the high-pressure oil, and the servo plunger 12 is displaced to the right by ⁇ x R ( ⁇ is The displacement amplification factor is determined by the internal lever mechanism. For a certain type of plunger pump, the value is fixed), thereby reducing the inclination of the plunger pump swashplate and reducing the displacement V of the plunger pump. Its constant power regulation characteristic curve corresponds to the BC section of curve (2) in FIG.
  • the right end of the inner spring 708 has an interaction force with the inner spring right contact seat 7102
  • the left end of the inner spring 708 has an interaction force with the inner spring left contact seat 7073
  • the flange 7074 has an interaction force with the stepped hole 7093.
  • the outer spring right contact seat 7091 has an interaction force with the right end of the outer spring 706, and the left end of the outer spring 706 has an interaction force with the outer spring adjustment screw 703. Under this condition, the two springs are simultaneously moved and compressed, and the connection manner is In series.
  • the series-parallel converter is in an operating state in which the inner spring 708 is connected in series with the outer spring 706, and the integrated stiffness coefficient k 0 is expressed as (k 1 is the stiffness coefficient of the inner spring and k 2 is the stiffness coefficient of the outer spring).
  • the compensating lever 801 is moved to the left, the inner spring 706 and the outer spring 708 are compressed, and the inner spring converter 707 and the outer spring converter 709 are also moved to the left.
  • the outer spring 706 is fitted with the outer spring adjusting screw 703, and the pre-tightening force of the serial-parallel converter can be changed by rotating the outer spring adjusting nut 704;
  • the inner spring converter 707 is fitted with the inner spring adjusting screw 701, and the nut is adjusted by rotating the inner spring 702 can change the spacing between the inner spring adjustment screw 701 and the inner spring converter 707, thereby adjusting the load pressure range of the series-parallel converter in series operation;
  • the structural size design of the spring seat 710 determines the outer spring converter 709 and the spring seat 710
  • the distance between the parallel extrusion surface 7103 and the parallel compression surface 7092 is one of the important dimensions that determine the pressure range of the parallel-parallel converter in parallel operation.
  • K B electromagnetic proportional pressure reducing input current - output pressure gain coefficient, when the solenoid proportional pressure reducing valve 9 hardware model is determined, this value can be regarded as a fixed value;
  • I the input current of the electromagnetic proportional pressure reducing valve 9
  • a 1 - the load pressure P 1 of the front pump 3 acts on the area on the compensation spool 802;
  • ⁇ the displacement amplification factor of the position following system, which is determined by the internal lever mechanism. For a certain type of plunger pump, the value is fixed;
  • K j displacement gradient coefficient of the plunger pump, which is determined by the size of the plunger pump structure and is fixed;
  • N the input speed of the plunger pump input shaft
  • the constant power regulation characteristic curve is composed of four straight lines, then the precise calculation and selection of the four linear lines are controlled. And the slope, the curve is softer than the constant power adjustment curve formed by the original three straight lines, and is closer to the constant power theoretical constant power hyperbola, as shown by the curve (2) and the curve (1) in Fig. 1, which can greatly improve
  • the degree of fitting between the constant power hyperbolic curves of the constant power theory makes the power matching between the plunger pump and the prime mover more reasonable and makes full use of energy.

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Abstract

一种双联轴向柱塞泵恒功率调节系统，包括主阀体（001）、外弹簧调节螺钉（703）、内弹簧调节螺钉（701）、外弹簧调节螺母（704）和内弹簧调节螺母（702），主阀体（001）内包括外弹簧（706）、内弹簧（708）、补偿杆（801）、补偿阀芯（802）及补偿活塞（805），主阀体内还包括内弹簧转换器（707）、外弹簧转换器（709）和弹簧座（710），内弹簧（708）位于内弹簧转换器（707）与弹簧座（710）之间，内弹簧转换器（707）贯穿外弹簧（706）及外弹簧转换器（709），外弹簧（706）的一端与外弹簧调节螺钉（703）接触、另一端与外弹簧转换器（709）接触，外弹簧转换器（709）与内弹簧转换器（707）相配合接触，内弹簧转换器（707）与内弹簧调节螺钉（701）、弹簧座（710）之间均留有间隙。该内外弹簧转换器可使两个弹簧在压缩过程中表现出四种弹簧刚度特性，使其表现出来的四种综合刚度系数在恒功率调节过程中，其特性曲线与理论恒功率双曲线的拟合程度更高；该机构可以大幅提高柱塞泵压力负载在15～25MPa范围工作时的能源利用率；有效的避免了柱塞泵和原动机工作在超负荷状态，提高硬件使用寿命。

Description

一种双联轴向柱塞泵恒功率调节系统及其应用 技术领域
本发明涉及一种双联轴向柱塞泵恒功率调节系统及其应用，属于液压传动部件技术领域。
背景技术
柱塞泵作为液压动力元件的主要成员，主要应用在高压、大流量、大功率的工况中；
在工程机械(如液压挖掘机)应用的液压系统中，由于一个工作周期内各个执行元件所需流量变化较大，为了达到控制方便、节约能源、减少波动和协同动作的目的，越来越多的选用多联泵，其中双联轴向柱塞泵是应用范围最广的一种。
双联轴向柱塞泵通常与原动机(如柴油机)配合使用，现在市场中常见的品牌型号，如川崎重工业株式会社设计制造的K3V系列柱塞泵，在工作状态下是利用恒功率调节系统保证柱塞泵充分利用原动机的输出功率，且避免原动机处于超负荷工作状态，达到节约能源、提高工作效率的目的；恒功率调节系统的工作原理是采集工作负载压力信号，与预设的并联弹簧进行力的平衡，将输出的位移信号放大处理，进而调节柱塞泵的输出排量大小；由于两个预设的并联弹簧自由长度不同、配合安装位置不同，从而使得其工作时分为均未压缩、单个压缩、并联压缩这三个阶段。
对应于上述三个工作阶段，柱塞泵的压力负载—输出流量曲线也分为三个阶段，如图1中曲线(3)所示，若柱塞泵中内弹簧、外弹簧的刚度系数分别为k1、k2，该并联弹簧组工作时表现出来的综合刚度系数为k0，那么AB段为弹簧均未压缩阶段(k0＝∞)，BD段为弹簧单个压缩阶段(k0＝k1)，DE段为两个弹簧并联压缩阶段(k0＝k1+k2)；直线的斜率大小与综合刚度系数k0直接相关；图1中曲线(1)是恒功率调节过程的理论恒功率双曲线，当实际的调节曲线越靠近该理论恒功率双曲线(1)时，称为拟合程度越高，那么液压泵对原动机输出功率的利用率越高，效率越高；
观察图1中曲线(1)、(3)可知，当压力负载在15～25MPa范围工作时，其恒功率特性曲线相对于理论曲线的拟合程度不高，能源浪费现象严重；理论上讲，弹簧数量越多，其该泵的恒功率调节特性曲线就有更多的分段，也就更贴近理论恒功率双曲线，得到最理想的调节特性。但是过多的弹簧数量会增加变量机构的设计难度和成本，复杂的结构更会导致系统稳定性降低，适得其反。
发明内容
针对现有技术的不足，本发明提供一种在不增加弹簧数量的基础上，使得其表现出更多弹簧刚度系数的双联轴向柱塞泵恒功率调节系统。
本发明还提供一种包含上述双联轴向柱塞泵恒功率调节系统的调节器。
本发明还提供一种包含上述调节器的双联轴向柱塞泵。
本发明的技术方案如下：
一种双联轴向柱塞泵恒功率调节系统，包括主阀体、外弹簧调节螺钉、内弹簧调节螺钉、外弹簧调节螺母和内弹簧调节螺母，主阀体内包括外弹簧、内弹簧、补偿杆、补偿阀芯及补偿活塞，在主阀体的两端设置有液压侧盖和弹簧侧盖，外弹簧调节螺钉通过弹簧侧盖安装在主阀体上并通过外弹簧调节螺母进行固定，内弹簧调节螺钉贯穿外弹簧调节螺钉并通过内弹簧调节螺母进行固定，其特征在于，主阀体内还包括内弹簧转换器、外弹簧转换器和弹簧座，内弹簧位于内弹簧转换器与弹簧座之间，内弹簧转换器贯穿外弹簧及外弹簧转换器，外弹簧的一端与外弹簧调节螺钉接触、另一端与外弹簧转换器接触，外弹簧转换器与内弹簧转换器相配合接触，内弹簧转换器与内弹簧调节螺钉、弹簧座之间均留有间隙。
优选的，所述内弹簧转换器的外形为圆柱体，圆柱体包含一空腔，内弹簧置于空腔内，内弹簧调节螺钉的一端贯穿内弹簧转换器及内弹簧。
优选的，所述外弹簧转换器为设有阶梯孔的圆环，所述内弹簧转换器的一端边缘处设有向外延伸的凸缘，当外弹簧转换器与内弹簧转换器相配合接触时，凸缘位于阶梯孔内。此设计的好处在于，通过凸缘与阶梯孔的接触配合，可使外弹簧转换器与内弹簧转换器配合更为紧密，配合精度更高。
优选的，所述弹簧座的一侧设有锥形凹槽，另一侧设有凸台，在凸台的外侧设有一圈可容纳凸缘的凹槽，凸台嵌入内弹簧，所述补偿杆的一端位于锥形凹槽内。
一种调节器，包括上述的恒功率调节系统、负流量反馈调节系统和阀控缸位置随动系统，所述负流量反馈调节系统包括负流量反馈油缸，所述阀控缸位置随动系统包括伺服阀、反馈杆、伺服柱塞和差动缸；补偿杆和负流量反馈油缸分别与伺服阀连接，用于控制伺服阀的阀芯运动，伺服柱塞设置在差动缸内并通过反馈杆与伺服阀连接，伺服柱塞通过反馈杆控制伺服阀的阀芯运动。
一种双联轴向柱塞泵，包括上述所述的调节器。
一种双联轴向柱塞泵的工作方法，包括以下步骤，
当柱塞泵的工作压力从空载开始逐渐升高，且未达到设定的柱塞泵工况功率时，柱 塞泵的出口压力随之增大，补偿阀芯上所受的力也逐渐增大，但此时补偿阀芯产生向左的作用力依然小于内弹簧和外弹簧向右的预紧力，补偿杆不产生位移，伺服阀和伺服柱塞均未移动，柱塞泵仍然以最大排量输出；
随着柱塞泵的工作压力继续升高，且达到了设定的柱塞泵工况功率时，柱塞泵的出口压力也随之增大，补偿阀芯上所受的力也逐渐增大，此时补偿阀芯产生向左的作用力克服内弹簧和外弹簧向右的预紧力，补偿杆向左产生位移并相应推动内弹簧转换器向左移动，补偿杆向左移动时带动伺服阀并通过伺服阀的作用使得伺服柱塞的大腔端接通高压油，伺服柱塞向右产生位移，进而减小柱塞泵斜盘倾角，减少柱塞泵的排量；
随着柱塞泵的工作压力继续升高，补偿杆继续向左移动，直至内弹簧转换器与内弹簧调节螺钉接触；
随着柱塞泵的工作压力继续升高，补偿杆继续向左移动，由于内弹簧转换器已经与固定的内弹簧调节螺钉接触，此时只有内弹簧被继续压缩；
随着柱塞泵的工作压力继续升高，补偿杆继续向左移动，当弹簧座和外弹簧转换器接触时，凸缘和阶梯孔逐渐脱离不再接触，外弹簧转换器单独向左移动，弹簧座同时压缩内弹簧和外弹簧；
当柱塞泵的工作压力降低，则柱塞泵的出口压力随之减小，其调节过程与柱塞泵工作压力升高时的调节过程相反。
本发明的有益效果在于：
本发明通过改进双联轴向柱塞泵内部的恒功率调节系统，在内弹簧和外弹簧处设计了一种全新的弹簧装配机构-串并联转换器(内弹簧转换器和外弹簧转换器)，该串并联转换器可使两个弹簧在压缩过程中表现出四种弹簧刚度特性，使其表现出来的四种综合刚度系数在恒功率调节过程中，其特性曲线与理论恒功率双曲线的拟合程度更高；该机构可以大幅提高柱塞泵压力负载在15～25MPa范围工作时的能源利用率；有效的避免了柱塞泵和原动机工作在超负荷状态，提高硬件使用寿命。
附图说明
图1为压力负载-输出流量特性曲线；
图2为调节器的原理图；
图3为恒功率调节系统的结构示意图；
图4为补偿杆的结构示意图；
图5为补偿阀芯的结构示意图；
图6a为弹簧座的主视图；
图6b为弹簧座的右视剖视图；
图7a为内弹簧转换器的主视图；
图7b为内弹簧转换器的左视剖视图；
图8a为外弹簧转换器的主视图；
图8b为外弹簧转换器的左视剖视图；
图9为内弹簧和外弹簧串联连接压缩时的状态图；
图10为内弹簧被压缩时的状态图；
图11为内弹簧被压缩结束/内弹簧和外弹簧并联连接开始压缩时的状态图；
图12为内弹簧和外弹簧并联连接压缩结束时的状态图；
图13为本发明中双联轴向柱塞泵的职能方块图；
其中：1、先导泵；2、后泵；3、前泵；4、原动机；5、差动缸；6、负流量反馈油缸；7、弹簧力补偿机构；8、液压力补偿机构；9、电磁比例减压阀；10、伺服阀；11、反馈杆；12、伺服柱塞；001、主阀体；002、液压侧盖；701、内弹簧调节螺钉；702、内弹簧调节螺母；703、外弹簧调节螺钉；704、外弹簧调节螺母；705、弹簧侧盖；706、外弹簧；707、内弹簧转换器；708、内弹簧；709、外弹簧转换器；710、弹簧座；801、补偿杆；802、补偿阀芯；803、补偿阀套；804、补偿堵塞；805、补偿活塞；8011、圆弧顶头；8012、环形凹槽；8013、补偿杆尾座；8021、补偿阀芯圆弧顶头；8022、本泵作用面；8023、他泵作用面；8024、补偿阀芯尾座；7101、凸台；7102、内弹簧右接触座；7103、并联挤压面；7104、锥形凹槽；7105、凹槽；7071、限位端面；7072、装配孔；7073、内弹簧左接触座；7074、凸缘；7091、外弹簧右接触座；7092、并联压缩面；7093、阶梯孔；
具体实施方式
下面通过实施例并结合附图对本发明做进一步说明，但不限于此。
实施例1：
本实施例提供一种双联轴向柱塞泵恒功率调节系统，该恒功率调节系统为柱塞泵上调节器内的恒功率调节系统，其具体结构如图3所示，在主阀体001内主要由弹簧力补偿机构7和液压力补偿机构8组成，其中弹簧力补偿机构7是由内弹簧调节螺钉701、内弹簧调节螺母702、外弹簧调节螺钉703、外弹簧调节螺母704、弹簧侧盖705、外弹簧706、内弹簧转换器707、内弹簧708、外弹簧转换器709、弹簧座710和密封圈等附 件构成的。液压力补偿机构8是由补偿杆801、补偿阀芯802、补偿阀套803、补偿堵塞804、补偿活塞805和密封圈等附件构成的。
在主阀体001的两端安装弹簧侧盖705和液压侧盖002，内弹簧708位于内弹簧转换器707与弹簧座710之间，内弹簧转换器707贯穿外弹簧706及外弹簧转换器709，外弹簧706的一端与弹簧侧盖705接触、另一端与外弹簧转换器709接触，外弹簧转换器709与内弹簧转换器707相配合接触，内弹簧转换器707与内弹簧调节螺钉701、弹簧座710之间均留有间隙。
外弹簧调节螺钉703由弹簧侧盖705安装在主阀体001上并通过外弹簧调节螺母704螺纹连接进行固定；内弹簧调节螺钉701为阶梯轴形状，其包括一个轴肩，内弹簧调节螺钉701的一端贯穿外弹簧调节螺钉703、内弹簧转换器707和内弹簧708，另一端通过内弹簧调节螺母702螺纹连接进行固定，内弹簧调节螺钉701的轴肩与内弹簧转换器707及外弹簧调节螺钉703之间均留有间隙，通过调节外弹簧螺钉和内弹簧螺钉在主阀体内的进入量来调节外弹簧和内弹簧的预紧力。
其中，内弹簧转换器的外形为圆柱体，如图7a、图7b所示，圆柱体包含一空腔和连通空腔的装配孔，内弹簧置于空腔内左侧的内弹簧左接触座7073上，内弹簧调节螺钉701的一端贯穿内弹簧转换器的装配孔7072后置于内弹簧708内。
外弹簧转换器709为设有阶梯孔的圆环，如图8a和图8b所示，圆柱体的一端边缘处设有向外延伸的凸缘7074，当外弹簧转换器709与内弹簧转换器707相配合接触时，凸缘7074位于阶梯孔7093内。通过凸缘与阶梯孔的接触配合，可使外弹簧转换器与内弹簧转换器配合更为紧密，配合精度更高。
如图6a、图6b所示，弹簧座710的一侧设有锥形凹槽7104，另一侧设有凸台7101，在凸台7101的外侧设有一圈可容纳凸缘7074的凹槽7105，凸台7101嵌入内弹簧708。补偿杆801的结构如图4所示，包括锥形的圆弧顶头8011、环形凹槽8012和补偿杆尾座8013，补偿杆左端的锥形圆弧顶头8011位于锥形凹槽7104内与其相配合，环形凹槽8012与阀控缸位置随动系统中伺服阀10的阀芯相连，右端的补偿杆尾座8013受到补偿阀芯802左端的补偿阀芯圆弧顶头8021的作用力。
如图6a和图6b所示，弹簧座710受到左端两个弹簧力的作用，其中凸台7101与内弹簧708的内圆配合，内弹簧右接触座7102与内弹簧的右端面配合接触，起到支撑内弹簧708的作用，并联挤压面7103与外弹簧转换器709的并联压缩面7092配合接触，其锥形凹槽7104与补偿杆的锥形圆弧顶头8011配合接触；
内弹簧708、外弹簧706、内弹簧转换器707和外弹簧转换器709共同组成了一个串并联转换器。内弹簧转换器707的限位端面7071与内弹簧调节螺钉701轴肩的距离决定了内弹簧708和外弹簧706处于串联工作状态的压力范围，装配孔7072与内弹簧调节螺钉701配合安装，内弹簧左接触座7073与内弹簧的左端面接触，凸缘7074与阶梯孔7093配合接触；
外弹簧转换器709的外弹簧右接触座7091和外弹簧706的右端面接触，并联压缩面7092与并联挤压面7103可以配合接触，阶梯孔7093与凸缘7074配合安装；
补偿阀芯802是一个阶梯环形结构，位于补偿阀套803内，其结构如图5所示，左端的补偿阀芯圆弧顶头8021与补偿杆尾座8013接触，中部的两个环形的本泵作用面8022和他泵作用面8023分别通有(双联轴向柱塞泵)前泵3和后泵2的出口压力P1、P2，右端的补偿阀芯尾座8024受到补偿活塞805的作用力，补偿活塞805通过外套补偿堵塞804置于补偿阀套803内，补偿阀套803位于主阀体001内。其中补偿活塞805的左端为连接油箱的低压油，右端的液压压力来自柱塞泵上的电磁比例减压阀9的电磁比例减压阀出口压力Pf。
本实施例提供的恒功率调节系统，在传统恒功率调节系统的基础上，设计了一种全新的弹簧装配机构-串并联转换器(内弹簧转换器和外弹簧转换器)，该串并联转换器可使两个弹簧在压缩过程中表现出四种弹簧刚度特性，使其表现出来的四种综合刚度系数在恒功率调节过程中，其特性曲线与理论恒功率双曲线的拟合程度更高；该机构可以大幅提高柱塞泵压力负载在15～25MPa范围工作时的能源利用率；有效的避免了柱塞泵和原动机工作在超负荷状态，提高硬件使用寿命。
实施例2：
本实施例提供一种双联轴向柱塞泵上所用的调节器，该调节器包括实施例1所述的双联轴向柱塞泵恒功率调节系统(由弹簧力补偿机构7、液压力补偿机构8组成)、一个负流量反馈调节系统(由负流量反馈油缸6组成)和一套阀控缸位置随动系统(主要由差动缸5、伺服阀10、反馈杆11、伺服柱塞12组成)。负流量反馈油缸6与阀控缸位置随动系统中的伺服阀10进行连接，用于控制伺服阀10的阀芯运动，恒功率调节系统通过其液压补偿机构8的补偿杆801与阀控缸位置随动系统中的伺服阀10进行连接，用补偿杆的运动来控制伺服阀10的阀芯运动，伺服柱塞12设置在差动缸5内并通过反馈杆11与伺服阀10连接，伺服柱塞12通过反馈杆11控制伺服阀10的阀芯运动，恒功率调节系统与负流量反馈系统之间无连接关系，同时恒功率调节系统与负流量反馈系统会通 过其他的控制方式保证只有一个在工作。
实施例3：
本实施例提供一种双联轴向柱塞泵，结构如图2所示，其包括先导泵1、主泵(前泵3和后泵2)、电磁比例减压阀9，在前泵3和后泵2上各安装一个如实施例2所述的调节器。
在实际的液压动力系统中，双联轴向柱塞泵通常与原动机连接，本实施例中原动机选为柴油机，原动机4通过输入轴拖动两个串联的柱塞通轴式后泵2、前泵3和安装在尾部的先导泵1，其中柱塞泵为工作执行元件提供高压动力油，先导齿轮泵为控制油路提供高压油。前泵3和后泵2分别配置了如实施例2所述的调节器1和调节器2。
电磁比例减压阀9可以随着输入电流i的变化，得到不同的电磁比例减压阀出口压力Pf，从而设定柱塞泵工况功率WC的大小。为了延长原动机的使用寿命，原则上规定柱塞泵工况功率WC应该小于原动机额定功率WM。
在实际的工程应用中，当柱塞泵的工作负载压力较小时，液压泵的实际输出功率WO未达到设定的工况功率WC，此时柱塞泵保持最大排量输出，以实现“轻载快速”的作业要求。随着负载压力的不断变大，液压泵的实际输出功率WO达到了设定的工况功率WC，若此时负载压力继续增大，则柱塞泵调节器会通过恒功率调节系统的作用，令柱塞泵的输出流量减小，从而避免原动机4超负荷工作，实现了“重载慢速”的自我保护工作模式。
恒功率调节系统将前泵3和后泵2的负载压力P1、P2作为控制信号，调节柱塞泵的输出排量，此种控制方式称为交叉功率控制策略，其优点是可以保证整个液压泵输出的总功率恒定，并且当其中一个泵所需功率减小时，另一泵可以自动利用剩余功率，充分利用原动机4功率。恒功率控制的目的是通过调节柱塞泵排量的方式，保证柱塞泵在工作过程中，既能与原动机4的输出功率之间建立良好的匹配关系，充分利用动力能源，又能够保证原动机4的输出功率不超过额定功率，延长机电液系统的使用寿命。
实施例4：
一种如实施例3所述的双联轴向柱塞泵的工作方法，包括以下步骤，
a)当柱塞泵的工作压力从空载开始逐渐升高，且未达到设定的柱塞泵工况功率WC时，前泵3、后泵2的出口压力P1、P2值也随之增大，补偿阀芯本泵作用面8022和他泵作用面8023上所受的力也逐渐增大，但此时补偿阀芯802产生向左的作用力依然小于内弹簧708和外弹簧706向右的预紧力，因此补偿杆801不产生位移，伺服阀10和伺服柱塞12均未移动，柱塞泵仍然以最大排量输出。其恒功率调节特性曲线对应于图1中曲线 (2)的AB段。
b)当柱塞泵的工作压力继续升高，且达到了设定的柱塞泵工况功率WC时，前泵3、后泵2的出口压力P1、P2值随之增大，本泵作用面8022和他泵作用面8023上所受的力也逐渐增大，此时补偿阀芯802产生向左的作用力克服内弹簧708和外弹簧706向右的预紧力，补偿杆801向左产生位移xR，阀控缸位置随动系统通过伺服阀10的作用，使得伺服柱塞12的大腔端接通高压油，伺服柱塞12向右产生位移μ·xR(μ为位移放大系数，该系数由内部杠杆机构决定，对于确定型号的柱塞泵，该值固定不变)，进而减小柱塞泵斜盘倾角，减少柱塞泵的排量V。其恒功率调节特性曲线对应于图1中曲线(2)的BC段。
该工况下内弹簧708的右端与内弹簧右接触座7102有相互作用力，内弹簧708的左端与内弹簧左接触座7073有相互作用力，凸缘7074与阶梯孔7093有相互作用力，外弹簧右接触座7091与外弹簧706的右端有相互作用力，外弹簧706的左端与外弹簧调节螺钉703有相互作用力，在该工况下两个弹簧同时被移动压缩，其连接方式是串联。此时，串并联转换器处于内弹簧708与外弹簧706串联的工作状态，表现出来的综合刚度系数k0大小为
Figure PCTCN2016084716-appb-000001
(k1为内弹簧的刚度系数，k2为外弹簧的刚度系数)。当补偿杆801向左运动时，内弹簧706和外弹簧708被压缩，内弹簧转换器707和外弹簧转换器709也向左运动。
c)当柱塞泵的工作压力继续升高，补偿杆801继续向左移动，直至内弹簧转换器707与内弹簧调节螺钉701接触，串联工作状态结束，此刻恒功率调节特性曲线对应于图1中曲线(2)的C点。
d)当柱塞泵的工作压力继续升高，补偿杆801继续向左移动，由于此时内弹簧转换器707已经与内弹簧调节螺钉701接触，此时外弹簧706停止压缩，只有内弹簧708继续压缩，该串并联转换器处于单个弹簧被压缩的工作状态，表现出来的综合刚度系数k0大小为k1。其恒功率调节特性曲线对应于图1中曲线(2)的CD段。
e)当柱塞泵的工作压力继续升高，补偿杆801继续向左移动，当并联挤压面7103和并联压缩面7092接触时，凸缘7074和阶梯孔7093逐渐脱离不再接触，外弹簧转换器709也向左移动，弹簧座710同时压缩内弹簧708和外弹簧706，此时串并联转换器处于内弹簧708和外弹簧706并联的工作状态。其恒功率调节特性曲线对应于图1中曲线(2)的DE段。
f)若柱塞泵的工作压力降低，则前泵3、后泵2的出口压力P1、P2值随之减小，调节原理相似，其调节过程与柱塞泵工作压力升高时的调节过程相反。
外弹簧706与外弹簧调节螺钉703配合安装，通过旋转外弹簧调节螺母704可以改变串并联转换器的预紧力；内弹簧转换器707与内弹簧调节螺钉701配合安装，通过旋转内弹簧调节螺母702可以改变内弹簧调节螺钉701与内弹簧转换器707的间距，进而调节串并联转换器处于串联工作状态的负载压力范围；弹簧座710的结构尺寸设计决定了外弹簧转换器709和弹簧座710之间的距离(即并联挤压面7103和并联压缩面7092之间的间距)，该距离是决定串并联转换器处于并联工作状态压力范围的重要尺寸之一。
建立该双联轴向柱塞泵的数学模型，根据变量机构的工作原理，可以绘制出恒功率调节机制的职能方块图，如图13所示。
其中：(Ⅰ)电磁比例减压阀9输出特性：
Pf＝KB·i
Pf—电磁比例减压阀9出口压力；
KB—电磁比例减压阀的输入电流-输出压力增益系数，当电磁比例减压阀9的硬件型号确定后，该值可视为固定值；
i—电磁比例减压阀9的输入电流；
(Ⅱ)恒功率调节系统力平衡方程：
Figure PCTCN2016084716-appb-000002
Af—电磁比例减压阀出口压力Pf作用于补偿活塞805上的面积；
A1—前泵3的负载压力P1作用于补偿阀芯802上的面积；
A2—后泵2的负载压力P2作用于补偿阀芯802上的面积；
k0—串并联转换装置的综合刚度系数；
k1—补偿杆801向左产生的位移；
k1—内弹簧708的刚度系数；
k2—外弹簧706的刚度系数；
xp—串并联转换装置的弹簧预紧长度；
(Ⅲ)阀控缸位置随动系统的位移传递关系：
xSF＝xmax-μ·xR
xSF—伺服柱塞12产生的位移；
xmax—伺服柱塞12的极限位移，当xSF＝xmax时，柱塞泵的排量达到最大值，该数值由柱塞泵结构尺寸决定，固定不变；
μ—位置随动系统的位移放大系数，该系数由内部杠杆机构决定，对于确定型号的柱塞泵，该值固定不变；
(Ⅳ)双联柱塞泵流量输出公式：
Q＝2·Kf·n·xSF
Q—双联柱塞泵的前泵3和后泵2输出总流量；
Kj—柱塞泵的排量梯度系数，该数值由柱塞泵结构尺寸决定，固定不变；
n—柱塞泵输入轴转速；
整理上述关系式，得到如下公式：
Figure PCTCN2016084716-appb-000003
假设前泵3和后泵2所受的负载压力P1、P2相等，即P1＝P2＝P0，当电磁比例减压阀9的输入电流i固定不变，原动机4以固定转速拖动柱塞泵时，n值大小不变，那么可以得到柱塞泵输出总流量Q与负载压力P0的关系如下：
Figure PCTCN2016084716-appb-000004
由上式可以看出，在柱塞泵的压力负载P0—输出流量Q特性曲线中，直线斜率与综合刚度系数呈负倒数关系。
在以根据上述串并联机构的四种工作状态可以分析出，弹簧的综合刚度系数k0如下：
Figure PCTCN2016084716-appb-000005
那么在带有弹簧串并联转换器的双联轴向柱塞泵恒功率调节系统中，其恒功率调节特性曲线由四条直线构成，那么经过精确计算和选型，控制四条直线的起始作用位置和斜率，该曲线比原有的三条直线构成的恒功率调节曲线更加柔和，更加逼近恒功率理论恒功率双曲线，如图1中的曲线(2)和曲线(1)，就可以大幅提高与恒功率理论恒功率双曲线之间的拟合程度，使得柱塞泵与原动机之间的功率匹配更加合理，充分利用能源。

Claims (7)

  1. 一种双联轴向柱塞泵恒功率调节系统，包括主阀体、外弹簧调节螺钉、内弹簧调节螺钉、外弹簧调节螺母和内弹簧调节螺母，主阀体内包括外弹簧、内弹簧、补偿杆、补偿阀芯及补偿活塞，在主阀体的两端设置有液压侧盖和弹簧侧盖，外弹簧调节螺钉通过弹簧侧盖安装在主阀体上并通过外弹簧调节螺母进行固定，内弹簧调节螺钉贯穿外弹簧调节螺钉并通过内弹簧调节螺母进行固定，其特征在于，主阀体内还包括内弹簧转换器、外弹簧转换器和弹簧座，内弹簧位于内弹簧转换器与弹簧座之间，内弹簧转换器贯穿外弹簧及外弹簧转换器，外弹簧的一端与外弹簧调节螺钉接触、另一端与外弹簧转换器接触，外弹簧转换器与内弹簧转换器相配合接触，内弹簧转换器与内弹簧调节螺钉、弹簧座之间均留有间隙。
  2. 如权利要求1所述的双联轴向柱塞泵恒功率调节系统，其特征在于，所述内弹簧转换器的外形为圆柱体，圆柱体包含一空腔，内弹簧置于空腔内，内弹簧调节螺钉的一端贯穿内弹簧转换器及内弹簧。
  3. 如权利要求2所述的双联轴向柱塞泵恒功率调节系统，其特征在于，所述外弹簧转换器为设有阶梯孔的圆环，所述内弹簧转换器的一端边缘处设有向外延伸的凸缘，当外弹簧转换器与内弹簧转换器相配合接触时，凸缘位于阶梯孔内。
  4. 如权利要求1所述的双联轴向柱塞泵恒功率调节系统，其特征在于，所述弹簧座的一侧设有锥形凹槽，另一侧设有凸台，在凸台的外侧设有一圈可容纳凸缘的凹槽，凸台嵌入内弹簧，所述补偿杆的一端位于锥形凹槽内。
  5. 一种调节器，包括权利要求1-4中任一项所述的恒功率调节系统、负流量反馈调节系统和阀控缸位置随动系统，所述负流量反馈调节系统包括负流量反馈油缸，所述阀控缸位置随动系统包括伺服阀、反馈杆、伺服柱塞和差动缸；补偿杆和负流量反馈油缸分别与伺服阀连接，用于控制伺服阀的阀芯运动，伺服柱塞设置在差动缸内并通过反馈杆与伺服阀连接，伺服柱塞通过反馈杆控制伺服阀的阀芯运动。
  6. 一种双联轴向柱塞泵，包括权利要求5中所述的调节器。
  7. 一种如权利要求6所述的双联轴向柱塞泵的工作方法，包括以下步骤，
    当柱塞泵的工作压力从空载开始逐渐升高，且未达到设定的柱塞泵工况功率时，柱塞泵的出口压力随之增大，补偿阀芯上所受的力也逐渐增大，但此时补偿阀芯产生向左的作用力依然小于内弹簧和外弹簧向右的预紧力，补偿杆不产生位移，伺服阀和伺服柱塞均未 移动，柱塞泵仍然以最大排量输出；
    随着柱塞泵的工作压力继续升高，且达到了设定的柱塞泵工况功率时，柱塞泵的出口压力也随之增大，补偿阀芯上所受的力也逐渐增大，此时补偿阀芯产生向左的作用力克服内弹簧和外弹簧向右的预紧力，补偿杆向左产生位移并相应推动内弹簧转换器向左移动，补偿杆向左移动时带动伺服阀并通过伺服阀的作用使得伺服柱塞的大腔端接通高压油，伺服柱塞向右产生位移，进而减小柱塞泵斜盘倾角，减少柱塞泵的排量；
    随着柱塞泵的工作压力继续升高，补偿杆继续向左移动，直至内弹簧转换器与内弹簧调节螺钉接触；
    随着柱塞泵的工作压力继续升高，补偿杆继续向左移动，由于内弹簧转换器已经与固定的内弹簧调节螺钉接触，此时只有内弹簧被继续压缩；
    随着柱塞泵的工作压力继续升高，补偿杆继续向左移动，当弹簧座和外弹簧转换器接触时，凸缘和阶梯孔逐渐脱离不再接触，外弹簧转换器单独向左移动，弹簧座同时压缩内弹簧和外弹簧；
    当柱塞泵的工作压力降低，则柱塞泵的出口压力随之减小，其调节过程与柱塞泵工作压力升高时的调节过程相反。
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AU2016343379B2 (en) 2018-12-20
CN105179221B (zh) 2017-01-11
CA2974845A1 (en) 2017-04-27
AU2016343379A1 (en) 2017-08-03

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