CN114876902A - Speed real-time tracking hydraulic control method and system and engineering machinery - Google Patents
Speed real-time tracking hydraulic control method and system and engineering machinery Download PDFInfo
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
- F15B13/00—Details of servomotor systems ; Valves for servomotor systems
- F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
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- B—PERFORMING OPERATIONS; TRANSPORTING
- B66—HOISTING; LIFTING; HAULING
- B66C—CRANES; LOAD-ENGAGING ELEMENTS OR DEVICES FOR CRANES, CAPSTANS, WINCHES, OR TACKLES
- B66C13/00—Other constructional features or details
- B66C13/18—Control systems or devices
- B66C13/20—Control systems or devices for non-electric drives
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- B—PERFORMING OPERATIONS; TRANSPORTING
- B66—HOISTING; LIFTING; HAULING
- B66C—CRANES; LOAD-ENGAGING ELEMENTS OR DEVICES FOR CRANES, CAPSTANS, WINCHES, OR TACKLES
- B66C23/00—Cranes comprising essentially a beam, boom, or triangular structure acting as a cantilever and mounted for translatory of swinging movements in vertical or horizontal planes or a combination of such movements, e.g. jib-cranes, derricks, tower cranes
- B66C23/62—Constructional features or details
- B66C23/82—Luffing gear
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- B—PERFORMING OPERATIONS; TRANSPORTING
- B66—HOISTING; LIFTING; HAULING
- B66D—CAPSTANS; WINCHES; TACKLES, e.g. PULLEY BLOCKS; HOISTS
- B66D1/00—Rope, cable, or chain winding mechanisms; Capstans
- B66D1/28—Other constructional details
- B66D1/40—Control devices
- B66D1/42—Control devices non-automatic
- B66D1/44—Control devices non-automatic pneumatic of hydraulic
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
- F15B13/00—Details of servomotor systems ; Valves for servomotor systems
- F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
- F15B13/04—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
- F15B13/044—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor operated by electrically-controlled means, e.g. solenoids, torque-motors
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
- F15B21/00—Common features of fluid actuator systems; Fluid-pressure actuator systems or details thereof, not covered by any other group of this subclass
- F15B21/02—Servomotor systems with programme control derived from a store or timing device; Control devices therefor
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
- F15B2211/00—Circuits for servomotor systems
- F15B2211/60—Circuit components or control therefor
- F15B2211/63—Electronic controllers
- F15B2211/6303—Electronic controllers using input signals
- F15B2211/6306—Electronic controllers using input signals representing a pressure
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
- F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
- F15B2211/00—Circuits for servomotor systems
- F15B2211/60—Circuit components or control therefor
- F15B2211/63—Electronic controllers
- F15B2211/6303—Electronic controllers using input signals
- F15B2211/6336—Electronic controllers using input signals representing a state of the output member, e.g. position, speed or acceleration
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- Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
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- Y02E—REDUCTION OF GREENHOUSE GAS [GHG] EMISSIONS, RELATED TO ENERGY GENERATION, TRANSMISSION OR DISTRIBUTION
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Abstract
The invention discloses a speed real-time tracking hydraulic control method, a control system and engineering machinery, wherein the speed real-time tracking hydraulic control method is used for adjusting the flow of a working medium of a hydraulic actuating element based on a flow valve, and comprises the following steps: acquiring the actual valve element overflowing opening area of the flow valve according to the instruction control speed and the real-time acquired valve front-valve back-valve differential pressure delta P of the flow valveA(ii) a Based on the actual demand valveCore over-current opening areaAAnd load feedback pressure P F Correcting the actually required valve core over-flow opening area A according to the corresponding relation, thereby obtaining a corresponding ideal valve core over-flow opening area Ad; and controlling the flow valve in real time by using a control signal corresponding to the ideal valve core overflowing opening area Ad. The invention can automatically realize the speed control of the hydraulic actuating element, not only improves the automation degree of the speed control, but also can avoid the interference of human factors to the control process, improve the control accuracy and improve the micro-motion stability.
Description
Technical Field
The invention relates to a speed control method of a hydraulic actuator, in particular to a speed real-time tracking hydraulic control method. Further, the invention relates to a corresponding speed real-time tracking hydraulic control system. In addition, the invention also relates to engineering machinery.
Background
The engineering machinery generally adopts a hydraulic actuator (typically a hydraulic cylinder or a hydraulic motor) to drive a working mechanism, and the smoothness of the movement of the working mechanism and the precision of fine movement under fine operation are of great importance to the working quality of the working mechanism. On some working mechanisms of engineering machinery, such as a lifting rope retracting mechanism driven by a hydraulic motor, a boom luffing mechanism of an automobile crane and the like, if the speed control of a hydraulic actuating element lacks smoothness, the safety of operation is seriously influenced, and even serious operation accidents are caused.
For example, referring to fig. 1, typically, the pilot operated directional control valve 2a is connected to the oil inlet path 1a and the oil return path (i.e., the oil path connected to the oil tank 2 a), and the pilot operated directional control valve 2a is connected to the rodless chamber of the luffing cylinder 5a via a first working oil path and connected to the rodless chamber of the luffing cylinder 5a via a second working oil path, and a pilot operated luffing balance valve 4a is provided on the first working oil path between the pilot operated directional control valve 2a and the rodless chamber and is driven by the pilot operated oil path 3a, so that the luffing cylinder 5a is driven to extend and retract by switching of the pilot operated directional control valve 2a, and the luffing cylinder 5a drives the luffing cylinder to luffing. However, the conventional amplitude-variable falling operation of the automobile crane boom is to control the opening degree of the valve core of the amplitude-variable balance valve 4a by providing stable pilot control oil from the outside, the rodless cavity of the amplitude-variable oil cylinder 5a has no pressure, and the boom realizes amplitude-variable falling by the self weight of the boom and the suspended load. In the amplitude falling process, the component force of the suspension arm acting on the oil cylinder is gradually increased, the oil pressure of the rodless cavity of the oil cylinder is gradually increased, and at the moment, if the opening degree of the valve core of the amplitude-changing balance valve 4a is kept unchanged, the amplitude falling speed of the suspension arm is faster and faster, and even the risk of stalling and falling is caused.
In order to avoid the risk of boom runout speed becoming fast or even stalling due to an increase in load, a hydraulic compensation method is usually used. Referring to fig. 1, in the hydraulic compensation, load pressure is fed back to a load pressure feedback port on the spool spring side of the amplitude-varying balance valve 4a through a feedback oil path, so that the spool opening pressure is properly increased, and when the load becomes large, the spool is gradually closed, so that the amplitude-varying speed is effectively maintained. However, the amplitude variation speed is completely controlled by a pilot oil pressure by a mechanic by experience to adjust the opening of the valve port of the main valve core, the speed cannot be accurately controlled, and the swinging is serious. When the traditional gravity amplitude-falling system is adopted to control the amplitude-changing falling speed, the following defects are provided: firstly, the speed in the amplitude-variable falling process cannot be accurately controlled, and the speed reduction control needs to be carried out by a manipulator through experience before stopping, otherwise, the suspension arm swings seriously, and the manipulator is easy to fatigue and generate risks easily in long-time operation. Secondly, the opening pressure of the valve core of the amplitude-variable balance valve is increased along with the increase of the load, so that the opening pressure of the amplitude-variable balance valve 4a is also increased due to the increase of the load in the amplitude-falling process of the suspension arm, the pressure of the pilot hydraulic control oil is also increased, and the control signal of the pilot handle for controlling the pressure of the pilot hydraulic control oil is gradually increased. However, in general, the minimum control signal of the pilot handle is set to a fixed value, which causes that when the load is low (such as when the boom has a large elevation angle), the minimum output control signal of the pilot handle is higher than the opening control pressure of the amplitude variation balance valve 4a, and the amplitude drop generates impact instantly; when the load is high (such as when the suspension arm is small-angle), the minimum output control signal of the handle is lower than the opening control pressure of the balance valve, the control dead zone of the handle is increased, the effective stroke is reduced, and the control precision of the variable amplitude balance valve is reduced. And thirdly, in essence, the amplitude falling speed of the traditional suspension arm amplitude changing system is adjusted by a mechanic through experience to adjust the input current of an electro proportional valve in a pilot control oil way and control the pilot oil pressure of an amplitude changing balance valve 4a to control the opening degree of a valve core, so that the flow passing through an oil cylinder is adjusted to achieve the purpose of changing the amplitude falling speed, and the speed is not easy to control in the amplitude falling process and has higher requirements on the mechanic.
In fact, in the application occasions of controlling the work stability of the hydraulic actuating element on the engineering machinery through the balance valve, the problems are common, but in view of the complexity of hydraulic control and the variability of working conditions, the defects cannot be effectively solved in the field, the smoothness of work is not ideal, and even a safe operation accident is caused.
In view of the above, a new speed control scheme for hydraulic actuators is needed.
Disclosure of Invention
The invention aims to solve the basic technical problem of providing a speed real-time tracking hydraulic control method which can automatically realize the speed control of a hydraulic actuating element in real time, not only improve the accuracy and the automatic real-time performance of the speed control, but also improve the control accuracy and the operation stability.
Further, the technical problem to be solved by the present invention is to provide a speed real-time tracking hydraulic control system, which can improve the accuracy and automatic real-time performance of speed control of a hydraulic actuator, and can improve the control accuracy and improve the operation stability.
In addition, the technical problem to be solved by the invention is to provide the engineering machinery, the amplitude variation speed of the suspension arm of the engineering machinery is more accurately controlled, automatic and real-time, the control is accurate, and the operation is stable.
In order to solve the above-mentioned technical problems,the invention provides a speed real-time tracking hydraulic control method which adjusts the flow of a working medium of a hydraulic actuator based on a flow valve, wherein the control method comprises the following steps: acquiring an actual valve element overflowing opening area A of the flow valve according to the instruction control speed and the real-time acquired valve front-valve and rear-valve differential pressure delta P of the flow valve; based on the actual required valve core over-flow opening area A and the load feedback pressure P F According to said load feedback pressure P F Correcting the actual valve core over-flow opening area A according to the corresponding relation, so that the actual valve core over-flow opening area A is corrected to be the ideal valve core over-flow opening area A d Thereby obtaining the corresponding ideal valve core overflowing opening area A d (ii) a The area A of the ideal valve core overflowing opening degree d The corresponding control signal controls the flow valve.
Specifically, the actual valve element overflowing opening area A is calculated by the following method:
converting the command control speed into a required flow Q according to the structural parameters of a working cavity of the hydraulic actuating element;
calculating the actual required valve core overflowing opening area A according to the following formula:
wherein: c d Is the orifice throttling constant of the flow passage of the flow valve; ρ is the hydraulic working medium density.
Preferably, the ideal valve core flow opening area A is calculated by the following formula d ：
Wherein: k is the load feedback pressure P F Influence coefficient on the valve core overflowing opening area.
More preferably, the flow valve is an electrically controlled flow valve, and the control signal is an electrical control signal when the ideal valve core is obtainedFlow opening area a d In the case of (2), the current parameter of the electrical control signal is calculated by the following formula:
wherein: k is a radical of 2 Is an ideal valve core over-flow opening area A under an electric control mode d A coefficient of influence correlation with the electrical control signal; i is c A control current for the flow valve; and a is a correction parameter in an electric control mode.
As a parallel preferred embodiment, the flow valve is a hydraulic control type flow valve, the control signal is a pressure control signal, and when the ideal valve core overflowing opening area a is obtained d In the case of (2), the pressure parameter of the control pressure signal is calculated by the following formula:
wherein: k is a radical of 3 Is an ideal valve core overflowing opening area A under the hydraulic control mode d A coefficient of influence correlation with the pressure control signal; p c Is the control pressure of the flow valve; b is the correction parameter in the hydraulic control mode.
Further preferably, the hydraulic control method further includes: and detecting the real-time speed of the hydraulic actuating element or a working mechanism driven by the hydraulic actuating element, comparing the real-time speed with the command control speed to obtain a speed difference value, and determining a compensation control signal parameter according to the speed difference value so as to compensate the control signal.
Specifically, the control signal is an electric control signal, and the compensation current is determined according to the speed difference value, and comprises at least one of the following items: in the case that the speed difference is positive, the compensation current I is adjusted b Determining as a positive value; in the case that the speed difference value is negative, the compensation current I is adjusted b Determining a negative value; when the speed difference is zero, the compensation current I is adjusted b Is determined to be zero.
Correspondingly to the above hydraulic control method of the present invention, the present invention provides a speed real-time tracking hydraulic control system, including a hydraulic control loop for driving a hydraulic actuator, the hydraulic control loop including a directional valve connected to the hydraulic actuator, the directional valve being connected to an oil inlet path and an oil return path, and an electrically controlled flow valve being connected to a working oil path between the directional valve and the hydraulic actuator, wherein the hydraulic control system further includes: the oil pressure sensors are respectively arranged on the working oil way parts on two sides of the flow valve; the controller is used for receiving command control speed signals and receiving pressure signals acquired by the oil pressure sensors in real time, comparing the pressure signals to obtain the pressure difference before and after the valve of the flow valve, obtaining the actual valve element overflowing opening area A of the flow valve according to the command control speed and the pressure difference delta P before and after the valve, and obtaining the actual valve element overflowing opening area A of the flow valve based on the actual valve element overflowing opening area A and the load feedback pressure P F According to said load feedback pressure P F Correcting the actual valve core over-flow opening area A according to the corresponding relation, so that the actual valve core over-flow opening area A is corrected to be the ideal valve core over-flow opening area A d Thereby obtaining the corresponding ideal valve core overflowing opening area A d And the area A of the valve core opening degree is equal to the ideal valve core opening degree d And the corresponding electric control signal controls the opening of the flow valve in real time.
Preferably, the flow valve is an electro-proportional flow control valve.
Specifically, the hydraulic actuator is a hydraulic cylinder, the flow valve is provided on a first working fluid line between a rodless chamber of the hydraulic cylinder and the selector valve, and a load pressure feedback port of the flow valve is fluidly connected to a portion of the first working fluid line between the rodless chamber and the flow valve so as to be able to introduce a load pressure to a load pressure feedback port on a spring chamber side of the flow valve.
As a specific parallel mode, the hydraulic actuator is a hydraulic motor, the flow valve is disposed in a first working oil path between a first working oil port of the hydraulic motor and the reversing valve, and a load pressure feedback port of the flow valve is fluidly connected to an oil path portion of the first working oil path between the first working oil port and the flow valve so as to be able to introduce a load pressure to a load pressure feedback port on a spring chamber side of the flow valve.
Preferably, the reversing valve is an electrically controlled reversing valve, and the electrically controlled reversing valve is electrically connected to the controller.
On the basis of the technical scheme of the speed real-time tracking hydraulic control system, the invention also provides an engineering machine, which comprises a suspension arm amplitude-variable hydraulic control system and a winch hydraulic control system, wherein the suspension arm amplitude-variable hydraulic control system is the speed real-time tracking hydraulic control system in any applicable technical scheme, and the hydraulic actuating element in the speed real-time tracking hydraulic control system is an amplitude-variable hydraulic cylinder; and/or the hoisting hydraulic control system is a speed real-time tracking hydraulic control system according to any one of the above applicable technical schemes, and the hydraulic execution element in the speed real-time tracking hydraulic control system is a hydraulic motor.
Preferably, the speed real-time tracking hydraulic control system further comprises a speed detector, the speed detector is used for detecting the real-time speed of the hydraulic actuating element or a working mechanism driven by the hydraulic actuating element, the speed detector is electrically connected to the controller, the controller is further used for comparing the real-time speed with the command control speed to obtain a speed difference value, and a compensation current is determined according to the speed difference value to compensate the electric control signal.
The present invention also provides an engineering machine, including: a processor and a memory storing computer program instructions; when the processor executes the computer program instructions, the speed real-time tracking hydraulic control method in any technical scheme is realized.
Through the technical scheme of the invention, the speed real-time tracking hydraulic control method and the hydraulic control system thereof have the following advantages: firstly, the degree of automation is high, the speed tracking precision is high, and the real-time performance of speed adjustment is excellent. The speed tracking control is based on a physical model of a hydraulic system, the real-time control concept comprehensively considers the influence factors of control signals, valve core displacement, flow area, the reaction of load feedback port pressure on the valve core displacement and the like, the command control speed and the front valve and the rear valve are used as input parameters, the corresponding control signals are used as output parameters, the speed tracking control is a typical two-input one-output nonlinear control system, and the speed tracking control is applied to the real-time debugging of relevant engineering machinery (such as an automobile crane amplitude-changing system) for regulating the flow of a hydraulic execution element by a load feedback type flow valve, so that the speed tracking precision is high, the real-time performance is excellent, and the engineering machinery works stably; secondly, the hydraulic control method and the hydraulic control system thereof have low cost, do not need a large amount of hardware equipment basically, are provided with corresponding sensors (such as oil pressure sensors) according to input parameters, and upgrade and reform the controller, so that the configuration on the existing engineering machinery can be effectively compatible, and the large-scale application is convenient to realize; thirdly, the hydraulic control method and the hydraulic control system thereof have wide application range and high practicability, can be compatible to be popularized and used on the engineering machinery host under all working conditions, and do not need to adjust other working condition parameters when being used under all working conditions. In addition, because the hydraulic control method and the hydraulic control system thereof have the advantages of obviously improving the smoothness of the hydraulic actuating element, the hydraulic control method and the hydraulic control system thereof can also additionally realize other functions, such as reducing the swing of the boom arm support in the amplitude falling process.
Further, in the preferred embodiment of the present invention, the hydraulic control method and the hydraulic control system thereof of the present invention adopt an open-loop + speed compensation control mode, and the speed compensation mode effectively avoids the influence of the adverse factors such as too large closed-loop control error, long stabilization time, and changes in working conditions, so that the real-time control progress is higher, and the work stability of the engineering machinery is more excellent.
Additional features and advantages of the invention will be set forth in the detailed description which follows.
Drawings
The following drawings are included to provide a further understanding of the invention and are incorporated in and constitute a part of this specification, illustrate embodiments of the invention and together with the description serve to explain the scope of the invention. In the drawings:
FIG. 1 is a hydraulic schematic diagram of a boom luffing hydraulic system of a prior art mobile crane;
FIG. 2 is a block diagram of the steps of a speed real-time tracking hydraulic control method of the basic embodiment of the present invention;
FIG. 3 is a hydraulic schematic diagram of an electro-proportional flow valve used in an embodiment of the present invention;
FIG. 4 is a hydraulic schematic diagram of a velocity real-time tracking hydraulic control system in accordance with an embodiment of the present invention, in which the hydraulic control circuit of the luffing hydraulic cylinder is taken as a representative example;
FIG. 5 is a schematic block diagram of the basic control concept of the speed real-time tracking hydraulic control method of the basic embodiment of the present invention;
FIG. 6 is a graph of a functional relationship between an area of flow and a control signal formed by an actual test according to an embodiment of the present invention, in which a fluctuation domain relationship graph is shown, which is formed on the basis of an ideal area of flow curve, due to the influence of a composite load feedback pressure on the area of flow, and due to the influence of actual operating condition factors such as oil temperature and viscosity;
fig. 7 is a schematic block diagram of a control signal compensation control concept of the hydraulic control method for real-time speed tracking in the preferred embodiment of the present invention.
Description of the reference numerals of the invention:
1, feeding an oil way; 2, a directional valve; 3, flow valve; a 3A load pressure feedback port; 4, a hydraulic cylinder; 4A rodless cavity; 5, a controller; 6 an oil pressure sensor; 7 oil tank.
Detailed Description
The following detailed description of the present invention is provided in conjunction with the accompanying drawings, and it is to be understood that the detailed description is provided for purposes of illustration and explanation and is not intended to limit the scope of the invention.
Since the core technical concept of the present invention itself has logic level progression and technical complexity, in order to help understanding of the present invention, first, a boom luffing hydraulic control system of an automobile crane is taken as an example, and the core technical concept of the present invention is described in detail in a manner of a specific application embodiment. On the basis of fully understanding the control logic and technical concept of the present invention through the exemplary application embodiment, since the control concept of the present invention is not limited to speed control of a luffing hydraulic cylinder, but can be generally applied to various hydraulic actuators (such as a hydraulic motor, etc.), and the control concept of the present invention is not limited to a current control manner, and can also be implemented by pressure control, etc., the speed real-time tracking hydraulic control method and the hydraulic control system thereof of the present invention will be further described from a more generally applicable level of the present invention.
Referring to fig. 3 and 4, in the exemplary embodiment of the speed real-time tracking hydraulic control method applied to the jib luffing system of the automobile crane, it can be seen that the main structure of the jib luffing system of the automobile crane is similar to that of the existing luffing hydraulic system, i.e. the main hydraulic arrangement structure of the telescopic hydraulic control circuit of the luffing hydraulic cylinder (i.e. the hydraulic actuator is the hydraulic cylinder 4) is unchanged, which is good in compatibility and applicability when the speed real-time tracking hydraulic control method of the invention is applied to the existing engineering machinery. Specifically, referring to fig. 4, the telescopic hydraulic control circuit of the luffing hydraulic cylinder comprises a directional valve 2 for directional switching control of the main oil path in telescopic mode. In the exemplary embodiment of fig. 4, the reversing valve 2 is an electrically controlled reversing valve, such as an electromagnetic reversing valve, but it is possible to use a hydraulically controlled reversing valve, a manual reversing valve, or an electro-hydraulic proportional reversing valve, which all can implement the reversing switching control. The reversing valve 2 generally adopts a three-position four-way reversing valve, an oil inlet of the reversing valve is connected to an oil inlet oil path 1, an oil return port of the reversing valve is connected to an oil tank 7 through an oil return path, a first working oil port is connected to a rodless cavity 4A of the variable amplitude hydraulic cylinder through a first working oil path, and a second working oil port is connected to a rod cavity of the variable amplitude hydraulic cylinder through a second working oil path. Thus, in normal telescopic control, when the reversing valve 2 is switched to the right position, the oil inlet oil path 1 feeds oil to the rodless cavity 4A of the variable amplitude hydraulic cylinder through the first working oil path, the rod cavity feeds oil to the oil tank 7 through the second working oil path and the oil return oil path, and the piston rod of the variable amplitude hydraulic cylinder extends out to drive the lifting arm to lift; when the reversing valve 2 is switched to the middle position, oil inlet and oil outlet of the amplitude-variable hydraulic cylinder are all cut off, and the suspension arm is in a locking state; when the reversing valve 2 is switched to the left position, the oil inlet oil circuit 1 supplies oil to the rod cavity of the amplitude-variable hydraulic cylinder through the second working oil circuit, meanwhile, the rodless cavity 4A of the amplitude-variable hydraulic cylinder returns oil to the oil tank 7 through the first working oil circuit and the oil return oil circuit, at the moment, the piston rod of the amplitude-variable hydraulic cylinder retracts to drive the suspension arm to fall in amplitude, and under the actual working condition, due to the dead weight of the suspension arm, the suspension arm compresses the piston rod of the amplitude-variable hydraulic cylinder to retract along with the inclination of the falling amplitude. As the suspension arm is very heavy, the amplitude of the suspension arm falls by gravity, and the load on the amplitude-variable oil cylinder is continuously increased along with the reduction of the angle of the crane arm, so that the amplitude falling speed is faster and faster, the controllability and the stability of the crane are influenced, and the working strength of an operator is increased. Under the condition, if the flow regulation and control are not carried out on the return oil of the rodless cavity 4A of the luffing hydraulic cylinder, the amplitude falling speed of the suspension arm can be accelerated, and even weightlessness and free dumping can occur, which is very dangerous. Although the conventional balance valve also adopts a load pressure feedback technology, and the flow regulation generally adopts electric control or hydraulic control, the regulation is limited by lack of effective and scientific control rules, and the current or the pressure is controlled artificially according to the feeling and experience, so that the speed control of the amplitude-variable hydraulic cylinder is unstable, the abnormal working conditions such as oscillation, pause and contusion exist, and even the working accidents occur under the condition of not-in-place control.
However, in the exemplary embodiment of the luffing system of a crane boom to which the speed real-time tracking hydraulic control method of the present invention is applied, as shown in fig. 4, a certain adaptation improvement is made in terms of hardware structure, specifically, a flow valve 3 is provided on the first working fluid path instead of the balance valve, and the flow valve 3 uses a load pressure feedback technology to participate in dynamic force balancing of a spool of the flow valve 3, and it should be understood that, although in the exemplary application embodiment of the present invention, as well as the basic embodiment and various preferred embodiments described below, the load pressure feedback directly refers to the load pressure in the working fluid path, within the technical concept of the present invention, it is not limited thereto, since the load feedback pressure can be easily detected in real time, it can also be simulated by using an electrically controlled or hydraulically controlled manner at one end of the spool of the flow valve, so as to participate in dynamic force balance of the valve core of the flow valve in the working process. The flow valve 3 in fig. 4 is an electric proportional flow control valve (specifically, an electro-hydraulic proportional flow control valve or an electromagnetic proportional flow control valve in general), which is a hydraulic element commonly used in the hydraulic field, but the electric proportional flow control valve adds load pressure feedback in the application of the present invention, specifically, the hydraulic schematic diagram of the electric proportional flow control valve is shown in fig. 3, the right side of the electric proportional flow control valve is an electric driving mechanism for driving a valve core to move to adjust the opening overflowing area of the valve core, the right side is a spring cavity side, and a load pressure feedback port 3A on the spring cavity side is used for connecting a load pressure feedback oil path. The electric proportional flow control valve is used as a core element for the flow control of the hydraulic control system, and the valve flow is not only related to a control signal (or control pressure) but also related to the pressure of the load pressure feedback port 3A, namely the load feedback port pressure and the control signal (or control pressure) jointly determine the flow passing through the valve. The valve can adopt various forms such as a proportional electromagnet direct drive type (namely, the proportional electromagnet outputs displacement to directly push the main valve core to move) or a proportional pressure reducing valve pressure drive type (namely, the pressure pushes the main valve core to move after the pressure is reduced), and the like. The electro-hydraulic proportional signal tends to make the valve core open greatly, and the pressure of the load feedback port tends to make the valve core close, i.e. the pressure of the load port of the main valve core is opposite to the control signal (or control pressure). It should be noted herein that although the exemplary embodiment of FIG. 4 employs a flow valve of an electronically controlled type, the control algorithm and control logic described below in the present invention is not limited to an electronically controlled type, as long as the control algorithm and control logic of the present invention is known, and either a hydraulically controlled type or a manually controlled type can be implemented.
The electric drive mechanism on the electric control side of the electric proportional flow control valve is electrically connected to a controller 5 for implementing control logic, while the load pressure feedback port 3A of the electric proportional flow control valve is connected to an oil path portion of the first working oil path between the rodless chamber 4A and the flow valve via a load pressure feedback oil path (shown by a dotted line) to enable introduction of a load pressure to the load pressure feedback port on the spring chamber side of the flow valve. The electric control side of the electric proportional flow control valve is connected to the oil path portion between the electric proportional flow control valve of the first working oil path and the reversing valve 2 through an oil drainage path shown by a dotted line, and the electric proportional flow control valve is mainly used for draining oil when oil leakage accumulated on the electric control side of the electric proportional flow control valve exists, so that the influence of the oil leakage accumulated on the electric control side on the accuracy of the load pressure feedback on the spring side when the boom falls is avoided.
Further, in terms of hardware configuration, the first hydraulic circuit is provided with hydraulic pressure sensors 6 on both sides of the electric proportional flow control valve, and the hydraulic pressure sensors 6 on both sides are connected to the controller 5, respectively. Meanwhile, the reversing valve 2 adopts an electric control reversing valve which is also connected to the controller 5 so as to facilitate telescopic reversing switching control.
On the basis of the hardware structure of the above exemplary embodiment, the following describes the speed real-time tracking hydraulic control method and the control rule and control logic of the control system thereof, referring to fig. 5, the control method of the present invention collects the differential pressure Δ P between the front valve and the rear valve of the electric proportional flow control valve and uses the differential pressure Δ P as a real-time feedback quantity, and uses the differential pressure Δ P and the command control speed signal as input quantities together, and integrates parameters such as the spool displacement of the flow valve 3, the flow area, the influence of the load feedback port pressure on the flow area, and the like into a controller (which can also be prestored as a parameter value corresponding table), so that the valve control current signal value corresponding to the theoretical speed can be calculated in real time, and the control method is a typical two-input one-output nonlinear control system.
Taking the amplitude-varying amplitude-falling control of the crane jib shown in fig. 3 as an example, the pressure of the rodless cavity 4A of the amplitude-varying hydraulic cylinder and the pressure behind the electric proportional flow control valve are collected, the collected data are transmitted to the controller 5 in real time, the current corresponding to the amplitude-falling speed to be controlled at a certain load can be calculated through the relational expression of the input current of the electric proportional flow control valve, the load feedback pressure and the amplitude-falling retraction speed of the amplitude-varying hydraulic cylinder, and then the current signal is input to the electric proportional flow control valve.
Specifically, the amplitude of the boom of the automobile crane falls, a piston rod of the amplitude-variable hydraulic cylinder is compressed by the self weight of the boom to fall, an oil inlet oil path does not actively supply oil to a rod cavity of the amplitude-variable hydraulic cylinder (the amplitude-variable hydraulic cylinder absorbs oil from the oil inlet oil path due to the retraction of the piston rod), the retraction speed of the amplitude-variable hydraulic cylinder is only related to the flow passing through an electric proportional flow control valve, and the angular speed of the amplitude of the boom of the automobile crane can be simply calculated by the retraction speed of the amplitude-variable hydraulic cylinder through a geometric relational expression (the end part of the piston rod of a general amplitude-variable hydraulic cylinder is hinged to the boom, the retraction speed of the piston rod at the hinged point can be regarded as the linear speed of the boom at the hinged point, and the angular speed of the amplitude of the boom can be easily converted through the linear speed of the hinged point and the distance (namely, the rotation radius) from the hinged point to the rotation point at the root of the boom.
The invention relates to a speed real-time tracking hydraulic control method and a control rule and a control logic realization form of a control system thereof.
Specifically, the command control speed is converted into the required flow Q according to the command control speed and the structural parameters of the working cavity of the amplitude-variable hydraulic cylinder. The commanded control speed is used as one of the input parameters, and the final realization is the boom lowering speed that the operator wishes to control, which can be directly the boom lowering speed, or the retraction speed of the luffing hydraulic cylinder, and as mentioned above, the boom lowering speed can be easily converted into the retraction speed of the luffing hydraulic cylinder according to the driving geometric relationship between the hydraulic actuator and the working mechanism. In the case that the command control speed is the speed of the working mechanism driven by the hydraulic actuator, for example, the angular speed of boom lowering of the crane truck can be simply calculated by the geometric relational expression from the retraction speed of the luffing cylinder (generally, the end of the piston rod of the luffing cylinder is hinged to the boom, and the retraction speed of the piston rod at the hinged point can be regarded as the linear speed of the boom at the hinged point, and the boom lowering angular speed can be easily converted by the linear speed of the hinged point and the distance (i.e., the rotation radius) from the hinged point to the rotation point at the root of the boom), and then, as for the rotation speed of the hoisting mechanism driven by the hydraulic motor, the rotation speed of the hoisting mechanism is multiplied by the drive reduction ratio to obtain the rotation speed of the hydraulic motor. In the case that the command control speed is the speed of the hydraulic actuator, according to the structural parameters of the working chamber of the luffing hydraulic cylinder, for the luffing hydraulic cylinder, the structural parameters of the working chamber herein may refer to the sectional area of the rodless chamber, and the retraction speed of the luffing hydraulic cylinder is multiplied by the sectional area of the rodless chamber 4A of the luffing hydraulic cylinder, so as to obtain the required flow rate Q of the hydraulic working medium required for achieving the command control speed (i.e. the retraction speed = required flow rate Q/rodless chamber sectional area), although it should be mentioned here in advance that, since the core concept of the present invention is not limited to be applied to the hydraulic cylinder, in the following more generally applicable basic embodiment, the hydraulic actuator of the present invention may cover the hydraulic actuator such as a hydraulic motor, and for the hydraulic motor, the structural parameters of the working chamber may be the sum of the volumes of the working chambers circumferentially arranged by the hydraulic motor (i.e. the rotation speed = required flow rate Q/circumferentially arranged working chambers arranged by the hydraulic motor) The sum of the volumes of). The working chamber configuration parameters of the hydraulic actuator are the necessary flow conversion basis, whether the commanded control speed is the speed of the hydraulic actuator or the speed of the working mechanism driven by the hydraulic actuator. On the basis, the actual required valve core overflowing opening area A is calculated according to the formula (1):
in formula (1): c d Is the orifice throttling constant of the flow passage of the flow valve (3), which can be checked by a known mechanical or hydraulic operating manualTo; a is the actual required valve core over-flow opening area of the electric proportional flow valve 3, and is a parameter obtained by calculation according to the formula (1), namely the valve core opening (namely the valve core over-flow opening area) actually required by the valve core of the electric proportional flow control valve for realizing the required flow Q; delta P is the pressure difference between the front valve and the rear valve of the electric proportional flow control valve, belongs to the second input parameter, and is obtained by comparing the pressures detected by the oil pressure sensors on the two sides of the electric proportional flow control valve; rho is the density of the hydraulic working medium, namely the density of oil.
From the above equation (1), the velocity (or flow) and orifice throttling constant C d The actual valve core overflowing opening area A, the pressure difference delta P between the front valve and the rear valve and the density rho of the hydraulic working medium are related, and the orifice throttling constant C d And the density rho of the hydraulic working medium is a value taken under an actual working condition, and the speed (or flow) is only related to the actual required valve core overflowing opening area A and the pressure difference delta P between the front valve and the rear valve.
Furthermore, the differential pressure delta P between the front valve and the rear valve is a real-time variable quantity, the core of the speed real-time tracking hydraulic control method and the hydraulic control system thereof is that the speed and the differential pressure delta P between the front valve and the rear valve are required to be controlled according to an instruction, the valve core overflowing opening area A is required to be automatically matched, and the corresponding control signal (or control pressure) is calculated in real time, wherein in the amplitude falling process of the suspension arm, the pressure behind the valve of the electric proportional flow control valve is communicated with the oil tank 7 and is approximately zero, so in the subsequent conversion and control strategy considering the influence factors of the actual working condition, the load feedback pressure P is used for realizing the load feedback F The valve before-valve differential pressure Δ P can be approximated equivalently.
Due to the complexity of flow control under actual working conditions, a large number of actual working condition tests are carried out in the actual research and development process of the project group, and a functional relation curve chart between the flow area and the control signal shown in fig. 6 is formed through a large number of actual working condition test data, wherein a fluctuation domain relation chart formed by the influence of the composite load feedback pressure on the flow area and the influence of actual working condition factors such as oil temperature, viscosity and the like on the basis of an ideal flow area curve is shown. As can be seen from fig. 6, the valve element flow opening area a and the control signal (e.g., electrical control signal or pressure control) are actually requiredSignal), i.e., the control signal tends to make the flow area large; the actual valve core over-flow opening area A and the load feedback pressure P are required F Negative correlation, i.e. load feedback pressure P F Tending to reduce the flow area. Considering the temperature, viscosity and load feedback port pressure P of oil F And the influence on the valve core flow area is equal, the valve core flow opening area A is actually required to be a region similar to the shadow area in FIG. 6 instead of a simple control curve similar to the ideal flow area, and the control algorithm is the process of solving the mathematical function of the region. How to reflect and satisfy the actual valve core overflowing opening area A by the clearly determined corresponding functional relationship between the ideal overflowing area (i.e. the ideal valve core overflowing opening area) and the control signal, i.e. by the control strategy with actual operability, is a difficult point of control logic.
In order to solve the difficulty of the control logic, the practical working condition test relation diagram of fig. 6 is properly referred, and the derivation process and relation of the following functions are theoretically defined in advance:
1) according to the above formula (1), it can be seen that the required flow Q is positively correlated with the actual required valve element flow opening area a, and according to fig. 6, the required flow Q can be obtained together with the electric control signal (or pressure control signal) and the load feedback pressure P F Correlation, which can be generally set to a nonlinear relationship, i.e., Q = f (I, P);
2) the hydraulic actuator speed V is linear with the required flow Q, i.e. V = k a ×Q，k a The proportional coefficient of the speed and the flow can be converted according to the displacement parameter of the working cavity of the hydraulic actuator as described above;
3) according to 1) and 2) above), the function of the hydraulic actuator speed V is V = k a Xf (I, P), namely the speed of the hydraulic actuator has a nonlinear function relation with the control current and the load pressure;
4) the relationship between the control electrical signal I (or pressure control signal) and the flow area a is analyzed as follows:
a) in the ideal state, assume the load feedback pressure P F Zero, i.e. the acting force of the pressure of the load feedback port at the acting surface of the spring cavity is zero, and no negative consideration is needed at the momentLoad feedback pressure P F The influence on the valve core opening flow area, the flow area a is only related to the electric control signal I (or the pressure control signal), i.e. a = f (I), which can be obtained by bench or real vehicle testing, and is generally a non-linear relationship, which can be given in the form of a piecewise function or an interpolation table, i.e. the solid curve at the lowest side in fig. 6, which represents the flow area as the ideal valve core flow opening area a d 。
b) Taking into account the load feedback pressure P F Influence on the flow area, load feedback pressure P F Acting in a spring cavity of the electric proportional flow control valve, the area of the valve core end surface exposed to the spring cavity is fixed, so that the acting force of the load feedback pressure is linearly related to the pressure, and the load feedback pressure P is accordingly F Linearly related to spool displacement, i.e. with load feedback pressure P F In fig. 6, the current flowing area a is shifted along the axis of the control current signal (or the control pressure), i.e. the area formed by the shaded area in fig. 6. In addition, the function domain relation curve of fig. 6 is formed based on a large amount of data of an actual working condition test, in the actual working condition, the influence of the temperature and viscosity of the hydraulic working medium (i.e. the temperature and viscosity of the oil, etc.), the upper boundary of the shadow area of fig. 6, and the lower boundary (i.e. the ideal valve core flow opening area a) d Curves) are not exactly translationally identical in shape, but there is some difference in curvature and shape fluctuation between the two.
Based on the above theoretical analysis, the further function calculation process is as follows:
first, a load feedback pressure P is established F And the area A of the ideal valve core overflowing opening degree d According to a relationship with said load feedback pressure P F Correcting the actual valve core over-flow opening area A according to the corresponding relation, so that the actual valve core over-flow opening area A is corrected to be the ideal valve core over-flow opening area A d Thereby obtaining the corresponding ideal valve core overflowing opening area A d 。
Specifically, as a practical condition summary correction formula, based on the actually measured function curve relationship of fig. 6, the current is input in proportion and the current is negativeThe opening area relation of the electric proportional flow control valve under the combined action of the load feedback pressure can be represented by an actual measurement empirical formula (2), so that the load feedback pressure P is established F Ideal valve core flow opening area A d And actually requiring the relationship between the valve core overflowing opening area A:
P F load feedback pressure for an electrically proportional flow control valve F Typically not zero. In actual condition, the load feedback pressure P F The non-zero criterion can be generally defined as > 25bar, although depending on the control accuracy requirements P can also be defined F For other pressure values (e.g. 15bar or 35 bar), the load feedback pressure is considered to be non-zero; k is the load feedback pressure P F An influence coefficient k on the valve element overcurrent opening area, which is obtained through actual test or calculation and is prestored, may be a linear relationship or a nonlinear relationship, and is shown in fig. 6; a. the d The ideal valve element overflowing opening area is defined as the valve element overflowing opening area corresponding to the change of the control signal under the condition that the flow valve 3 is in the no-load feedback pressure state, and it needs to be explained in advance here that in the actual working condition, because the hydraulic working medium always exists in the hydraulic pipeline under the actual working condition, the load feedback pressure P is the load feedback pressure P F It is generally difficult to be absolutely zero, and P is usually defaulted F Under the condition of less than or equal to 25bar, the load feedback pressure P F Is zero, i.e. no load feedback pressure is defaulted, of course, P can be defined according to the requirement of control precision F At other pressure values ≦ e.g. 15bar or 35bar, the load feedback pressure P F Is zero; a is the actual required spool flow opening area a, which has been obtained according to the above equation (1). It is additionally stated here that the following is about the load feedback pressure P F The criterion of zero or not can be referred to the description here, and the description is not repeated.
It is further explained here that, for the person skilled in the art, load reversals are concernedPressure feed P F And the influence coefficient k on the valve core overflowing opening area is related to the control ratio of the electric proportional flow control valve. k represents the influence of the control pressure of the load port on the flow area when the current of the electric proportional flow control valve is fixed.
The specific calculation and testing procedure for the k value can be exemplified as follows: a fixed control current value (or control pressure value), such as 550mA (or 15 bar), is supplied to the electric proportional flow control valve 3, and according to the small-hole overflow formula (1), the pressure P is applied to the load port F Is a certain value P F =P F1 The flow area is A = A1, the load port pressure is artificially increased, P F Is a certain value P F =P F2 When the flow area a = a2, k = (a 2-a 1)/(P) F2 - P F1 ）。
During the actual test, the aforementioned process may be repeated for another fixed control current value (or control pressure value), such as 570mA (or 18 bar), for the electric proportional flow control valve 3; theoretically, if the current values are taken sufficiently large, k can be fitted to a curve. See the calculated numerical examples illustrated in the following table:
the load feedback pressure P is mainly solved in this step F Influence on the flow area, i.e. load feedback pressure P F The influence relationship of the area domain of FIG. 6 is substantially to solve A = f (I, P) F ) To a problem of (a).
Secondly, according to the obtained ideal valve core overflowing opening area A d Calculating the magnitude of the input current
The load feedback pressure P is already solved in the last step F The influence on the flow area is not considered any more when the feedback pressure P of the load port is not taken into consideration F In the hydraulic control system for the amplitude variation of the suspension arm of the automobile crane in the electric control type shown in fig. 4, the control signal is the electric control signal, and it should be noted that, within the scope of the technical idea of the present invention, the PWM control signal (analog signal) and the CAN bus signal are commonly used in the actual working condition because the electric control signal of the present inventionThe control signal encompasses what is commonly referred to in the art as an electrical control signal, including a strong electrical control signal, a weak electrical control signal, and the like. On the basis, only the ideal valve core overflowing opening area A needs to be solved at the moment d The relation to the input current, i.e. the relation a = f (i), is as follows:
wherein: k is a radical of 2 Is an ideal valve core over-flow opening area A under an electric control mode d The influence correlation coefficient between the valve core and the electric control signal is specifically the valve core overcurrent opening area A formed in practical test d The ideal valve core overcurrent opening area A corresponding to different values in a relation function curve taking a vertical coordinate and a current control signal as a horizontal coordinate d The slope of the tangent line at the point of (a). Reference may be made here to fig. 6, but it is noted that fig. 6 is the ideal valve spool flow opening area a at the lowermost side d Based on the influence factors of the composite actual working condition, wherein the solid curve at the lowest side is the ideal valve core flow opening area A d A function curve with the control signal; i is c A control current for the flow valve; a is a correction parameter in an electric control mode, and specifically represents the area A of the ideal valve core overflowing opening degree d The intercept on the abscissa axis of the curve of the said relationship function with the control signal.
Thus, the control current I required by the electric proportional flow control valve can be obtained by gradually calculating based on the two input parameters c Therefore, the electric proportional flow control valve can be controlled in real time, the valve element overflowing opening area A is actually needed, the needed flow Q matched with the command control speed is obtained, and the hydraulic actuating element and the working mechanism driven by the hydraulic actuating element are controlled to run at the needed command control speed in real time, for example, the boom dropping amplitude speed is not unstable in working, and even stalling occurs to operation safety accidents.
It should be additionally noted that, in the above formula of the present invention, the function curve relationship is mainly formed based on actual tests, and therefore, the related calculation formula mainly uses the geometric trigonometric function relationship related to the function curve and the correction of the actual condition factor, and so on to perform calculation, so that more attention is paid to the matching and calculation of the numerical value, and no matching is needed for the unit of the parameter, which is common in the control formula of the engineering machine, and the following description will not be given in addition.
As described above, although the specific control process is described by taking a boom luffing hydraulic control system of a typical automobile crane as an example in order to help understanding of the technical concept of the present invention, it can be seen that the essence of the core concept of the speed real-time tracking hydraulic control method and the hydraulic control system thereof of the present invention is: the differential pressure delta P before and after the valve is a real-time acquisition quantity, the differential pressure delta P and the command control speed are jointly used as input quantities, after a core algorithm (contents of a core kernel of the control method, a derivation process of a mathematical function and the like), a control current signal I (or control pressure) can be automatically given or given according to a prestored data corresponding table, the overflowing area is dynamically adjusted along with the change of load feedback pressure (namely the overflowing area of a valve core is dynamically adjusted along with the difference of amplitude angle, load and the like and the differential pressure delta P after the valve before and after the valve is different), the overflowing area A can be matched in a shadow area of a graph 6, and under the condition of electric control or hydraulic pressure control, the whole process does not need to intervene according to experience and feeling of an operator. Along with the change of the command control speed command, the pressure difference delta P between the front valve and the rear valve and the like, the current control signal (or the control pressure) can be automatically calculated, the hydraulic control system can automatically follow the speed curve, and the tracking precision is high. Taking the amplitude-variable falling of the automobile crane as an example, the physical quantities such as the amplitude-variable falling angle of the boom arm support, the crane load and the like are all converted into the control variables (such as pressure, flow, overflowing area and the like) of the hydraulic system, wherein the influence of the amplitude-variable angle, the load and the like on the system is fed back by the load to feed back the pressure P F The method is embodied by considering the influences of actual working condition factors such as a hydraulic system control current signal (or control pressure), a valve core overflowing area (the characteristics of a valve), oil temperature and the like, and is suitable for full-working-condition floor application.
However, it should be noted that, although the above-mentioned hydraulic control system for luffing of a boom of an automobile crane is described as an example, it is mainly described in connection with an exemplary embodiment for facilitating technical understanding, the control concept of the present invention is not limited to speed control of a luffing hydraulic cylinder, but can be generally applied to various hydraulic actuators (such as a hydraulic motor, etc.), and the control concept of the present invention is not limited to a current control manner, and can also be implemented by pressure control, etc., and control hardware thereof is not limited to a controller, etc., since the control formulas of the steps of the present invention are already clarified, and relevant input parameters and corresponding output control parameters of actual tests have corresponding relationships and can be prestored as an actual measurement parameter corresponding table, in case of allowing appropriate discontinuous control, the technical idea of the invention can be basically realized by adopting manual control.
For this reason, the basic embodiment and various preferred embodiments of the speed real-time tracking hydraulic control method and the hydraulic control system thereof according to the present invention will be described in a more general aspect, and in the description, various implementation forms and modification forms to which the present invention can be applied will be additionally described, and these implementation forms and modification forms should be considered to fall within the scope of the present invention within the technical idea of the present invention.
Referring to fig. 2 and 5, a speed real-time tracking hydraulic control method according to a basic embodiment of the present invention adjusts a flow rate of a working medium of a hydraulic actuator based on a flow valve 3, wherein the control method includes the steps of: acquiring an actual required valve element overflowing opening area A of the flow valve 3 according to the instruction control speed and the real-time acquired valve front-valve back-valve differential pressure delta P of the flow valve 3; based on actual demand valve core overflowing opening area A and load feedback pressure P F According to said load feedback pressure P F Correcting the actual valve core over-flow opening area A according to the corresponding relation, so that the actual valve core over-flow opening area A is corrected to be the ideal valve core over-flow opening area A d Thereby obtaining the corresponding ideal valve core overflowing opening area A d (ii) a The ideal valve core overflowing opening area A d Corresponding control signal real-time controlThe flow valve 3.
In the basic embodiment, it is particularly emphasized that the valve element flow opening area a is required to pass through the load feedback pressure P F The corresponding relation is corrected to obtain the ideal valve core overflowing opening area A d The correspondence may be a linear correspondence of an actual condition test, or may be a calculation formula (2) summarized below as a preferred actual condition, or may be a correspondence recorded by a table after the actual condition test, which all fall within the technical concept of the present invention, and implementation variations that can be thought of by those skilled in the art are suggested by the technical concept of the present invention.
For the basic implementation of the speed real-time tracking hydraulic control method of the present invention, it is applicable to any situation where the load feedback type flow valve 3 is used to adjust the working medium flow of the hydraulic actuator, and any hydraulic control system, no matter its main function or basic function, only needs its whole or some local loop to use the load feedback type flow valve, and uses the speed real-time tracking hydraulic control method of the present invention to control the flow valve 3 to adjust the working medium flow of the hydraulic actuator, and thus it belongs to the protection scope of the present invention. To this end, it should be understood that the flow valve of the present invention employs a load feedback technique, so that the load pressure participates in the dynamic force balance of the valve core of the flow valve, which may be generally referred to as a "load feedback flow valve", and one end of the flow valve may be a hydraulic control type valve core driving mechanism, an electric control type (e.g., electromagnetic proportion, electro-hydraulic proportion, etc.) valve core driving mechanism or a manual type valve core driving mechanism, and the other end of the flow valve may be a spring cavity side, and the spring cavity is provided with a load pressure feedback port, and the load pressure feedback port is connected to the working oil path of the hydraulic actuator in the hydraulic system through a feedback oil path, so that the load pressure of the hydraulic actuator can be introduced to the spring cavity side. The whole loop or the local loop of various hydraulic systems only adopts the load feedback type flow valve, and the load feedback type flow valve belongs to a control object which can be applied by the speed real-time tracking hydraulic control method; of course, within the scope of the technical idea of the present invention, without being limited thereto, since the load feedback pressure can be easily detected in real time, it can also participate in dynamic force balance of the flow valve spool during operation by simulating the load feedback pressure at one end of the flow valve spool in an electrically or hydraulically controlled manner. In addition, the hydraulic actuator may be various actuators in the hydraulic field, such as a hydraulic cylinder, a hydraulic motor, and the like, and the working medium of the hydraulic actuator is generally referred to as hydraulic oil.
On the basis of the basic embodiment, in terms of the implementation form of obtaining the actual required valve element overflowing opening area a of the flow valve 3 according to the command control speed and the pressure difference Δ P before and after the valve, preferably, as described in the typical application embodiment of the luffing hydraulic system of the crane boom of the automobile, the calculation can be effectively and quickly performed through the flow calculation formula of the orifice, that is, the command control speed is converted into the required flow Q according to the structural parameters of the working chamber of the hydraulic actuator; then, the actual required valve core overflowing opening area A is calculated according to the following formula:
wherein: c d Is the orifice throttling constant of the flow passage of the flow valve 3; a is the actual required valve core overflowing opening area of the flow valve 3; delta P is the pressure difference between the front valve and the rear valve of the flow valve 3; ρ is the hydraulic working medium density.
However, the embodiment of obtaining the actual valve element flow opening area a of the flow valve 3 according to the command control speed and the valve front-valve and valve rear-valve differential pressure Δ P is not limited to the above preferred embodiment according to the formula (1), which can be obtained in various forms, for example, since the commanded control speed can be easily converted to the desired flow Q of the hydraulic actuator, under the condition of obtaining the required flow Q, the simplest method for obtaining the actual required valve core overflowing opening area A under the actual working condition with load pressure is to test the actual working condition, thereby forming a data table corresponding to the required flow Q and the actual required valve core overflowing opening area A under different load pressures of the flow valve 3 or the pressure difference between the front valve and the rear valve, under the condition of required flow Q and conveniently detected load pressure or valve front-valve back-pressure difference, a controller or an operator can effectively and correspondingly find the corresponding actual required valve core overflowing opening area A according to a data table. Therefore, the embodiment of obtaining the actual valve element flow opening area a of the flow valve 3 according to the command control speed and the valve front-valve and valve back-valve differential pressure Δ P, which can be considered by those skilled in the art in the light of the present disclosure, should fall within the technical idea of the present invention.
Further, on the basis of obtaining the actual required valve core overflowing opening area A, the invention obtains the ideal valve core overflowing opening area A d The realization form of (1) needs to consider the influence factors of the actual working condition, especially the load pressure, for this reason, the actual working condition test is needed, a domain function relation similar to the graph 6 is formed through the actual working condition data, and the load feedback pressure P is obtained from the domain function relation F And (4) pre-storing an influence coefficient k on the valve core overflowing opening area. For this reason, as a specific embodiment, in the second step of the above-described basic embodiment, it is preferable that the ideal spool flow opening area a is calculated by the following formula d ：
Wherein: p F Is the load feedback pressure of the flow valve (3), the load feedback pressure P F Should not be zero; k is the load feedback pressure P F The influence coefficient on the valve core over-flow opening area can be obtained through actual tests and is prestored; a. the d The valve element is an ideal valve element overflowing opening area, and is defined as the valve element overflowing opening area corresponding to the change of the control signal under the condition that the flow valve 3 is in a no-load feedback pressure state. It should be additionally noted that, as described above, the formula (2) is only a more preferred embodiment of the present invention, and the ideal valve core flow opening area a is obtained by correcting the valve core flow opening area a through the corresponding relationship d The method can have various implementation forms, and the corresponding relation can be a linear corresponding relation of actual working condition tests or an actual working conditionAnd after the test, the corresponding relation recorded by the table is passed. In the formula (2) of the preferred embodiment, k is the load feedback pressure P F The influence coefficient on the valve core overflowing opening area can be considered as a corresponding relation, and k is related to the load feedback pressure P F The product of the two is a corrected value of the actual required valve core flow opening area A in the basic embodiment.
It should be particularly noted that, under the condition of continuous real-time control, the control pattern of the present invention may adopt an electric control pattern or a hydraulic control pattern (that is, the control signal may be a pressure control signal or a current control signal), and in the working condition function relationship diagram formed by the actual test in fig. 6, the working condition function relationship diagram is an actual working condition test diagram implemented by adopting the electric control pattern based on the jib luffing system of the automobile crane. However, the principle of the actual working condition test is similar under the mode of adopting hydraulic control, and the load feedback pressure P under the corresponding control mode can be obtained F And (3) calculating the influence coefficient of the valve core overflowing opening area by adopting a formula (2). It is within the scope of the present invention to employ any control scheme, as long as it employs the technical principles based on fig. 6 and equation (2).
Further, based on the preferred implementation form of the electric control shown in fig. 6, if the flow valve 3 is an electric control type flow valve, the control signal is an electric control signal, and in the second step of the basic embodiment of the hydraulic control method of the present invention, preferably, when the ideal valve core flow opening area a is obtained d The current parameter of the electrical control signal may then be calculated by the following formula:
wherein: as described above, k 2 Is an ideal valve core over-flow opening area A under an electric control mode d Influence correlation coefficient between the valve core and the electric control signal is specifically an ideal valve core overcurrent opening area A corresponding to different values in a relation function curve formed according to actual test d In the curve of the relationship function, withIdeal valve core flow opening area A d The control signal is used as a vertical coordinate and the control signal is used as a horizontal coordinate; i is c Is the control current of the flow valve 3; a is a correction parameter in an electric control mode, and specifically represents the area A of the ideal valve core overflowing opening degree d The intercept on the abscissa axis of the curve of the said relationship function with the control signal.
It should be noted that the applicability of the principle of the above formula (3 a) is general, for example, the same applies to the pressure control signal, and with the formula (3 a), the control ratio of the electric proportional balance valve is known to those skilled in the art, and the control current I and the control pressure P are then c Satisfy I = k × P c + compensation factor (where k is as explained above in the technical meaning), so that it is Ad = k in the case of a pilot signal with pilot pressure 2 *P c The principle of the calculation of + a is equally applicable, except that in actual testing, k is 2 And a may vary in magnitude, but its principle of forming a functional relationship curve like the actual test in fig. 6 has general applicability. Therefore, the invention provides a parallel specific preferred embodiment, the flow valve 3 can be a flow valve of a hydraulic control type, the control signal is a pressure control signal, and when the ideal valve core overflowing opening area A is obtained d In the case of (2), the pressure parameter of the pressure control signal is calculated by the following formula:
wherein: k is a radical of 3 The method is an influence correlation coefficient between the ideal valve core over-flow opening area Ad and a pressure control signal in a hydraulic control mode, and specifically is the ideal valve core over-flow opening area A corresponding to different values in a relation function curve formed according to actual test d The slope of the tangent line at the point (a) in the curve of the relationship function as the area of the desired valve element opening d The control signal is used as a vertical coordinate and the control signal is used as a horizontal coordinate; p c Is the control pressure of the flow valve 3; b is the area A of the ideal valve core opening degree of overflowing in the hydraulic control mode d A coefficient of influence between control signals, which is specifically indicative of the ideal valve elementOpen area of flow A d The intercept on the abscissa axis of the curve of the said relationship function with the control signal.
It should be noted that, based on the certainty and definite uniqueness of the control, the core technical concept of the present invention is that after the actual working condition influencing factors (such as load pressure, oil temperature, oil viscosity, etc.) are considered step by step and the parameters are processed, the ideal valve core flow opening area a is still corresponded to d This is the ideal valve core flow opening area A d The solid curve has a clear correspondence with the control signal, i.e., the solid curve on the lowest side in fig. 6, i.e., the solid curve with the overcurrent area as the ordinate and the control current as the abscissa. Under the condition of adopting hydraulic control, the hydraulic control pressure is needed as an abscissa in the actual test, the control and adjustment principles are similar, but for different control types, the influence factors of actual working conditions are influenced, and under different control types, a relation function curve shows the ideal valve core overflowing opening area A d The intercept on the ordinate axis of (a) is influenced in a controlled manner, with a certain difference in value, but this as a whole still falls within the technical idea of the present invention.
As a more preferable embodiment, the speed real-time tracking hydraulic control method of the present invention may further introduce speed compensation, and such speed compensation may be effectively applied to various embodiments of the present invention, including the luffing hydraulic control system of the mobile crane according to the above exemplary application example. Specifically, referring to fig. 7, the hydraulic control method of the present invention further includes: and detecting the real-time speed of the hydraulic actuating element or a working mechanism driven by the hydraulic actuating element, comparing the real-time speed with the command control speed to obtain a speed difference value, and determining a compensation control signal parameter according to the speed difference value so as to compensate the control signal.
For example, in the case where the control signal is an electrical control signal, the current control signal is compensated by detecting the difference between the commanded control speed and the fed back real-time speed. I.e. determining a compensation current based on the speed difference, including at least one of: in the case that the speed difference value is a positive number, the step of comparing the speed difference value with the reference valueCompensating current I b Determining as a positive value; in the case that the speed difference value is negative, the compensation current I is adjusted b Determining a negative value; when the speed difference is zero, the compensation current I is adjusted b Is determined to be zero. The method comprises the following specific steps: when the difference value between the command control speed and the real-time speed is a positive number, compensating the current I b Is a positive value; when the difference value between the command control speed and the real-time speed is negative, the compensation current I b Is a negative value; when the difference between the control command speed and the real-time speed is zero, the compensation current I b Is zero (i.e., no compensation for current flow).
On the basis of the above embodiments of the speed real-time tracking hydraulic control method and the typical embodiments of the boom luffing system applied to the automobile crane, the tower crane or other engineering machinery, the invention also provides a speed real-time tracking hydraulic control system as a typical hydraulic telescopic or forward and reverse rotation control loop generally applied to a hydraulic actuating element, which comprises a hydraulic control loop for driving the hydraulic actuating element, wherein the hydraulic control loop comprises a reversing valve 2 connected to the hydraulic actuating element, the reversing valve 2 is connected to an oil inlet path 1 and an oil return path, and an electrically controlled flow valve 3 is connected to a working oil path between the reversing valve 2 and the hydraulic actuating element, wherein the hydraulic control system further comprises: an oil pressure sensor 6, the oil pressure sensor 6 being provided on the working oil path portions on both sides of the flow valve 3, respectively; and the controller 5 is electrically connected with the flow valve 3 and each oil pressure sensor 6, the controller 5 is used for receiving an instruction control speed signal and receiving pressure signals acquired by each oil pressure sensor 6 in real time, comparing the pressure signals to obtain a pressure difference delta P before and after the valve of the flow valve 3, obtaining an actual valve element overflowing opening area A of the flow valve 3 according to the instruction control speed and the pressure difference delta P before and after the valve, and obtaining the actual valve element overflowing opening area A based on the actual valve element overflowing opening area A and the load feedback pressure P F According to said load feedback pressure P F The corresponding relation corrects the actually required valve core over-flow opening area A, so that the actually required valve core over-flow opening area A is corrected to be the ideal valve core over-flow opening area A d Thereby obtaining the corresponding ideal valve core overflowingOpening area A d And the ideal valve core overflowing opening area A d The corresponding control signal controls the opening of the flow valve 3 in real time.
It should be noted that, as mentioned above, the technical concept of the present invention is not limited to the typical control circuit of the hydraulic actuator, and any partial portion of the hydraulic system adopting the control concept of the present invention should fall within the protection scope of the present invention. For the speed real-time tracking hydraulic control system, a hydraulic cylinder is adopted as a hydraulic actuating element, so that a telescopic control system of the hydraulic cylinder can be formed, for example, the luffing system of the suspension arm of the automobile crane is a typical application example, and the hydraulic actuating element is a concrete engineering machinery application embodiment of the hydraulic cylinder; if the hydraulic actuator adopts a hydraulic motor, a forward and reverse rotation control system of the hydraulic motor is formed.
In order to facilitate and speed control, the speed real-time tracking hydraulic control system of the basic embodiment adopts an electric control type and the controller 5 is used for specific control, and regarding each of the above-mentioned disclosures and corresponding control steps of the present invention, the speed real-time tracking hydraulic control method is the same as or corresponding to the speed real-time tracking hydraulic control method, and is only specifically executed by the controller 5 in the hydraulic control system, so that the implementation hardware related to the related preferred embodiment will not be repeated or correspondingly described in detail in each embodiment of the hydraulic control system below, and will be described only in relation to the implementation hardware related to the related preferred embodiment.
Preferably, in the hydraulic control system for real-time speed tracking of the basic embodiment, the flow valve 3 of the electrically controlled type may be an electro-proportional flow control valve, and specifically may be an electro-hydraulic proportional flow control valve or an electromagnetic proportional flow control valve.
As a typical hydraulic setting form suitable for boom lowering conditions or other work mechanism lowering conditions with increased weight, in order to avoid the risk of stalling, the flow valve 3 generally needs to be arranged on the working oil path to which the rodless chamber of the hydraulic cylinder is connected, specifically, the hydraulic actuator may be the hydraulic cylinder 4, the flow valve 3 is arranged on the first working oil path between the rodless chamber 4A of the hydraulic cylinder and the directional valve 2, and the load pressure feedback port 3A of the flow valve 3 is fluidly connected to the oil path portion of the first working oil path between the rodless chamber 4A and the flow valve 3 so as to be able to introduce the load pressure to the load pressure feedback port 3A on the spring chamber side of the flow valve, in which case the introduced load pressure is opposite to the force of the drive mechanism on the other side of the flow valve 3 on the spool, to a certain extent dynamically balancing the spool position.
Of course, for the hoisting hydraulic control system, the hydraulic actuator may be a hydraulic motor, a flow valve 3 may be disposed on the first working oil path between the first working oil port of the hydraulic motor and the reversing valve 2 (in this embodiment, the first working oil port of the hydraulic motor is used for driving the hydraulic motor to reverse through the switching of the reversing valve), and a load pressure feedback port 3A of the flow valve 3 is fluidly connected to an oil path portion of the first working oil path between the first working oil port and the flow valve 3, so as to be able to introduce load pressure to a load pressure feedback port 3A on the spring cavity side of the flow valve.
As for the direction change valve 2, which is an oil path switching element for the working direction of the hydraulic actuator, as described above, various types may be adopted, and preferably, in order to facilitate the centralized control by the controller 5, the direction change valve 2 may be an electrically controlled direction change valve, and the electrically controlled direction change valve is electrically connected to the controller 5.
In the preferred technical concept of the above speed compensation, the hydraulic control system needs to be provided with a speed detector for detecting the real-time speed of the hydraulic actuator or the working mechanism driven by the hydraulic actuator, the speed detector is electrically connected to the controller 5, so that the controller 5 is also used for comparing the real-time speed with the command control speed to obtain a speed difference value, and determining a compensation current according to the speed difference value so as to compensate the control signal. Such a speed detector can detect the movement speed of the working mechanism driven by the hydraulic actuator, for example, a rotation angle sensor can be arranged for the boom falling width speed; more generally, the working speed of the hydraulic actuator can be directly detected, for example, the telescopic speed of the hydraulic cylinder, and a cylinder wire displacement sensor can be typically adopted. Such a speed detector may take various suitable forms within the technical idea of the present invention as long as it can detect a real-time speed.
In addition, on the basis of the technical scheme of the speed real-time tracking hydraulic control system, the invention also provides an engineering machine, which comprises a boom amplitude hydraulic control system, wherein the boom amplitude hydraulic control system is the speed real-time tracking hydraulic control system, and the hydraulic actuating element in the speed real-time tracking hydraulic control system is an amplitude hydraulic cylinder.
In practical application, the engineering machine may be an automobile crane or a tower crane.
Alternatively, the hoisting hydraulic control system of the construction machine may also adopt the speed real-time tracking hydraulic control system, and the hydraulic actuator in the speed real-time tracking hydraulic control system is a hydraulic motor.
Further, the present invention is also a construction machine, wherein the construction machine comprises: a processor and a memory storing computer program instructions; and when the processor executes the computer program instructions, the speed real-time tracking hydraulic control method is realized.
As can be seen from the above description of the speed real-time tracking hydraulic control method and the hydraulic control system of the invention, the key technical concept of the invention comprises: firstly, the invention controls the speed and the pressure difference delta P behind the valve of the flow valve according to the instruction (the oil tank pressure is considered to be zero under the ideal working condition, and the pressure difference delta P behind the valve can be approximately equivalent to the load feedback pressure P F ) Real-time matching of the valve element over-flow opening area A is actually required and converted into the ideal valve element over-flow opening area A d So as to calculate the current control signal (or control pressure) needed to be adopted in real time; secondly, the influence of the actual working condition on the flow area is integrated, in the test of the actual working condition, the measured data fully considers the influence of the temperature, the viscosity, the load feedback port = pressure and the like of the oil liquid on the flow area of the valve core, and the flow area under the actual working condition is not a simple control curve but a similar control curveLike a domain of the shaded area in fig. 6, converting the domain function into a definite and clearly controllable control strategy is the core of the control method of the present invention; third, although the electric proportional flow control valve is a valve element which is used in a hydraulic field, the present invention introduces load pressure feedback into the electric proportional flow control valve, so that the flow rate of the valve is not only related to the current control signal (or control pressure), but also related to = = pressure of a load feedback port, that is, the load feedback pressure and the current control signal (or control pressure) jointly determine the flow rate flowing through the valve. In the electric control type hydraulic control method, in view of the convenience of electric control, the core of the invention is to calculate the relationship among the current control signal, the flow area and the influence of load feedback pressure on the flow area, as shown in the shaded area of fig. 6; fourthly, in a preferred mode, the hydraulic control method further introduces real-time speed compensation, real-time errors of pure open-loop control can be well solved through the speed compensation, and the hydraulic control method is convenient to apply to all working conditions.
In summary, the speed real-time tracking hydraulic control method and the hydraulic control system thereof have the following advantages: firstly, the degree of automation is high, the speed tracking precision is high, and the real-time performance of speed adjustment is excellent. The speed tracking control is based on a physical model of a hydraulic system, and the real-time control concept comprehensively considers control signals, valve core displacement, overflowing area and load feedback port pressure P F For influencing factors such as the reaction of valve core displacement, the command control speed and the front valve and the rear valve of the valve are used as input parameters, and the corresponding control signal is used as an output parameter, so that the system is a typical two-input one-output nonlinear control system, and is applied to the real-vehicle debugging on the relevant engineering machinery (such as an automobile crane amplitude-changing system) for regulating the flow of a hydraulic execution element by a load feedback type flow valve, the speed tracking precision is high, the real-time performance is excellent, and the engineering machinery works stably; secondly, the hydraulic control method and the hydraulic control system thereof have lower cost, do not need a large amount of hardware equipment basically, are provided with corresponding sensors (such as oil pressure sensors) according to input parameters, and upgrade and reform the controller, so that the configuration on the existing engineering machinery can be effectively compatible, and the large-scale realization of the configuration on the large-scale engineering machinery is facilitatedApplication; thirdly, the hydraulic control method and the hydraulic control system thereof have wide application range and high practicability, can be compatible to be popularized and used on the engineering machinery host under all working conditions, and do not need to adjust other working condition parameters when being used under all working conditions. In addition, because the hydraulic control method and the hydraulic control system thereof have the advantages of obviously improving the smoothness of the hydraulic actuating element, the hydraulic control method and the hydraulic control system thereof can also additionally realize other functions, such as reducing the swing of the boom arm support in the amplitude falling process. The speed real-time tracking hydraulic control method can be effectively applied to hydraulic systems with high requirements on stability and jogging property, such as variable amplitude falling of an automobile crane. Under the electric control mode of the controller, except the input or selection of the command control speed, the automatic control of the amplitude falling speed can be realized basically without human participation, the automation degree of the amplitude falling control of the crane is improved, assistance is provided for the unmanned and intelligent hoisting machinery, the interference of human factors to the control process can be avoided, the control accuracy is improved, and the micro-motion stability is improved.
Furthermore, the hydraulic control method and the hydraulic control system thereof also adopt a control mode of open loop plus speed compensation, and the speed compensation mode effectively avoids the influence of the disadvantages of overlarge closed-loop control error, longer stabilization time, change of working conditions and the like, so that the real-time control progress is higher, and the working stability of the engineering machinery is more excellent.
The preferred embodiments of the present invention have been described in detail with reference to the accompanying drawings, however, the present invention is not limited to the specific details of the above embodiments, and various simple modifications can be made to the technical solution of the present invention within the technical idea of the present invention, and these simple modifications are within the protective scope of the present invention.
It should be noted that the various features described in the above embodiments may be combined in any suitable manner without departing from the scope of the invention. The invention is not described in detail in order to avoid unnecessary repetition.
In addition, any combination of the various embodiments of the present invention is also possible, and the same should be considered as the disclosure of the present invention as long as it does not depart from the spirit of the present invention.
Claims (15)
1. A speed real-time tracking hydraulic control method is characterized in that the control method comprises the following steps:
acquiring an actual required valve element overflowing opening area A of the flow valve (3) according to the instruction control speed and the real-time acquired valve front-valve and rear-valve differential pressure delta P of the flow valve (3);
based on the actual required valve core over-flow opening area A and the load feedback pressure P F According to said load feedback pressure P F Correcting the actual valve core over-flow opening area A according to the corresponding relation, so that the actual valve core over-flow opening area A is corrected to be the ideal valve core over-flow opening area A d Thereby obtaining the corresponding ideal valve core overflowing opening area A d ；
The area A of the ideal valve core overflowing opening degree d The corresponding control signal controls the flow valve (3) in real time.
2. The hydraulic control method for real-time speed tracking according to claim 1, wherein the actual valve element flow opening area A is calculated by:
converting the command control speed into a required flow Q according to the structural parameters of a working cavity of the hydraulic actuating element;
calculating the actual required valve core overflowing opening area A according to the following formula:
wherein: c d Is the orifice throttling constant of the flow passage of the flow valve (3); ρ is the hydraulic working medium density.
3. The hydraulic control method for real-time speed tracking according to claim 1, wherein the ideal spool flow opening area A is calculated by the following formula d ：
Wherein: k is the load feedback pressure P F And influence coefficient on the valve core overflowing opening area.
4. The hydraulic control method for real-time speed tracking according to claim 1, characterized in that the flow valve (3) is an electrically controlled flow valve, the control signal is an electrical control signal, and the ideal valve core flow opening area A is obtained when the ideal valve core flow opening area A is obtained d In the case of (2), the current parameter of the electrical control signal is calculated by the following formula:
wherein: k is a radical of 2 The method is an influence correlation coefficient between the ideal valve core overflowing opening area Ad and an electric control signal in an electric control mode; i is c Is the control current of the flow valve (3); and a is a correction parameter in an electric control mode.
5. The hydraulic control method for real-time speed tracking according to claim 1, characterized in that the flow valve (3) is a hydraulic control type flow valve, the control signal is a pressure control signal, and the ideal valve core flow opening area A is obtained after the control signal is a pressure control signal d In the case of (2), the pressure parameter of the pressure control signal is calculated by the following formula:
wherein: k is a radical of 3 For ideal valve under hydraulic control modeCore over-current opening area A d A coefficient of influence correlation with the pressure control signal; p c Is the control pressure of the flow valve (3); b is the correction parameter in the hydraulic control mode.
6. The hydraulic control method for real-time tracking of speed according to any one of claims 1 to 5, further comprising:
and detecting the real-time speed of the hydraulic actuating element or a working mechanism driven by the hydraulic actuating element, comparing the real-time speed with the command control speed to obtain a speed difference value, and determining a compensation control signal parameter according to the speed difference value so as to compensate the control signal.
7. The method of claim 6, wherein the control signal is an electrical control signal, and wherein determining a compensation current based on the speed difference comprises at least one of:
in the case that the speed difference is positive, the compensation current I is adjusted b Determining as a positive value;
in the case that the speed difference value is negative, the compensation current I is adjusted b Determining a negative value;
when the speed difference is zero, the compensation current I is adjusted b Is determined to be zero.
8. The utility model provides a speed real-time tracking hydraulic control system, includes the hydraulic control circuit that is used for driving hydraulic actuator, and this hydraulic control circuit includes connect in hydraulic actuator's switching-over valve (2), and this switching-over valve (2) are connected in oil feed oil circuit (1) and oil return oil circuit, and are connected with flow valve (3) of automatically controlled type on this switching-over valve (2) and the working oil circuit between the hydraulic actuator, its characterized in that, hydraulic control system still includes:
the oil pressure sensors (6), the oil pressure sensors (6) are respectively arranged on the working oil path parts on the two sides of the flow valve (3); and
controller (5)The controller (5) is electrically connected to the flow valve (3) and the oil pressure sensors (6), the controller (5) is used for receiving command control speed signals and receiving pressure signals acquired by the oil pressure sensors (6) in real time, comparing the pressure signals to obtain a pressure difference delta P before and after the valve of the flow valve, obtaining an actual valve element overflowing opening area A of the flow valve (3) according to the command control speed and the pressure difference delta P before and after the valve, and obtaining the actual valve element overflowing opening area A and a load feedback pressure P based on the actual valve element overflowing opening area A and the load feedback pressure P F According to said load feedback pressure P F Correcting the actual valve core over-flow opening area A according to the corresponding relation, so that the actual valve core over-flow opening area A is corrected to be the ideal valve core over-flow opening area A d Thereby obtaining the corresponding ideal valve core overflowing opening area A d And the area A of the valve core opening degree is equal to the ideal valve core opening degree d And the corresponding electric control signal controls the opening of the flow valve in real time.
9. The speed real-time tracking hydraulic control system according to claim 8, characterized in that the flow valve (3) is an electro-proportional flow control valve.
10. The speed real-time tracking hydraulic control system according to claim 9, characterized in that the hydraulic actuator is a hydraulic cylinder (4), the flow valve (3) is provided on a first working fluid path between a rodless chamber (4A) of the hydraulic cylinder and the directional control valve (2), and a load pressure feedback port (3A) of the flow valve (3) is fluidly connected to a fluid path portion of the first working fluid path between the rodless chamber (4A) and the flow valve (3) to enable introduction of a load pressure to the load pressure feedback port (3A) on a spring chamber side of the flow valve.
11. The speed real-time tracking hydraulic control system according to claim 9, wherein the hydraulic actuator is a hydraulic motor, the flow valve (3) is provided on a first working oil path between a first working oil port of the hydraulic motor and the directional control valve (2), and a load pressure feedback port (3A) of the flow valve (3) is hydraulically connected to an oil path portion of the first working oil path between the first working oil port and the flow valve (3) to enable introduction of a load pressure to the load pressure feedback port (3A) on a spring chamber side of the flow valve.
12. The speed real-time tracking hydraulic control system according to claim 8, wherein the reversing valve (2) is an electrically controlled reversing valve electrically connected to the controller.
13. An engineering machine, which comprises a boom amplitude hydraulic control system and a winch hydraulic control system, and is characterized in that the boom amplitude hydraulic control system is a speed real-time tracking hydraulic control system according to any one of claims 8 to 10 and 12, and the hydraulic actuator in the speed real-time tracking hydraulic control system is an amplitude hydraulic cylinder; and/or
The hoisting hydraulic control system is a speed real-time tracking hydraulic control system according to any one of claims 8 to 9, 11 and 12, and the hydraulic actuator in the speed real-time tracking hydraulic control system is a hydraulic motor.
14. A working machine according to claim 13, wherein the speed real-time tracking hydraulic control system further comprises a speed detector for detecting a real-time speed of the hydraulic actuator or a working mechanism driven by the hydraulic actuator, the speed detector being electrically connected to the controller (5), the controller being further adapted to compare the real-time speed with the commanded control speed to obtain a speed difference, and to determine a compensation current based on the speed difference to compensate the electrical control signal.
15. A work machine, characterized in that the work machine comprises: a processor and a memory storing computer program instructions;
the processor, when executing the computer program instructions, implements a speed real-time tracking hydraulic control method as claimed in any one of claims 1-7.
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| CN202210808877.8A CN114876902A (en) | 2022-07-11 | 2022-07-11 | Speed real-time tracking hydraulic control method and system and engineering machinery |
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| CN202210808877.8A CN114876902A (en) | 2022-07-11 | 2022-07-11 | Speed real-time tracking hydraulic control method and system and engineering machinery |
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| CN117450122A (en) * | 2023-12-25 | 2024-01-26 | 中联重科土方机械有限公司 | Hydraulic system, control method and rescue equipment |
| WO2024124799A1 (en) * | 2022-12-14 | 2024-06-20 | 中联重科股份有限公司 | Adaptive control method, apparatus and device |
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