CN204041615U - Controlled variable cross section oil hydraulic cylinder and hydraulic control system thereof - Google Patents

Abstract

A controllable variable-section hydraulic cylinder and its hydraulic control system include a cylinder barrel, a first-stage piston rod, and a second-stage piston rod. The first-stage piston rod is sealed and installed on the cylinder barrel through the piston ring and the end cover, the first-stage piston rod is the cylinder barrel of the second-stage piston rod, and the second-stage piston rod is installed in the first-stage piston rod through the piston ring and the end cover seal. The control system of the above-mentioned variable cross-section hydraulic cylinder selects the primary piston rodless cavity, the primary piston rod cavity, the secondary piston rodless cavity, and the secondary piston rod cavity by controlling the two-position three-way electromagnetic switch valve and the three-position four-way servo valve. Whether the rod cavity is connected to the oil outlet of the pump, the oil tank is connected or cut off, so as to obtain different effective area. A load force sensor is arranged on the secondary piston rod to measure the load on the secondary piston rod in real time, that is, the load of the variable-section hydraulic cylinder. According to the load feedback, through the control of the electromagnetic switch valve group and the servo valve group, select the appropriate effective area to realize the matching of the maximum output force of the hydraulic cylinder and the load force, and realize the output flow of the pump and the load flow through the adaptive variable function of the variable pump. The matching, so as to realize the matching of the output power of the variable pump and the load power, so as to improve the system efficiency.

Description

Controllable variable section hydraulic cylinder and its hydraulic control system

technical field

The utility model relates to a controllable variable section hydraulic cylinder and a hydraulic control system thereof.

Background technique

In recent years, mobile robots have made great progress in the fields of structure, perception, path planning, and control, and have realized complex functions such as entertainment services, human-computer interaction, and extreme environment detection. limited, which restricts the practical application of mobile robots. Existing studies have shown that when the output pressure of the power source exceeds 3.5MPa, the hydraulic drive system with the same power has a higher power density than the pure electromechanical drive system. Adopting hydraulic drive system is an effective way to improve load capacity. At present, many research institutes are beginning to try to use hydraulic power systems to drive mobile robots, such as BIGDOG and petman developed by Boston Dynamics, KenKen II hydraulically driven quadruped robots of the Italian Polytechnic University, and high-performance robots funded by China's 863 High-tech Research and Development Program. The quadruped robot project explicitly calls for a hydraulic drive system.

Due to weight and volume limitations, the hydraulic drive system of mobile robots adopts a single pump source-multi-actuator system structure. The efficiency of this type of hydraulic drive system is very low. The main reason is that the load of each actuator is different at the same time, and the load of the same actuator is also different at different times. One pump source cannot simultaneously provide power to the loads of multiple actuators. Matching, generally choose a high-power actuator load for matching, resulting in a large amount of throttling loss in other actuator branches, resulting in low efficiency.

Inefficiency will cause the following problems: the power requirement of the power source is high, and the weight and volume of the power source will increase; the energy (such as gasoline) required to complete the same work will increase, and the weight will increase; the performance index requirements of the hydraulic components of the system will increase, and the hydraulic components will The weight and volume of the cooling system will increase; the heating of the system will be more serious, the power of the cooling system will become larger, and the volume and weight of the cooling system will increase. Therefore, the inefficiency will seriously affect the payload capacity of the mobile robot.

There are many methods to improve the efficiency of single-pump source-multi-actuator hydraulic drive system, such as independent throttling control of inlet and outlet oil, electro-hydraulic hybrid power and energy recovery technology, load-sensitive pump control technology, hydraulic transformer, etc. These technologies have limited energy-saving effects, and do not consider the volume and weight of the system, so it is difficult to use them on mobile robots.

According to the load, the effective area of the hydraulic cylinder is changed in real time, so that the load pressure of each actuator branch is close to the output pressure of the pump source, and the relationship between the output flow of the pump and the load flow of each branch is automatically adjusted through the variable adaptive mechanism of the pump. And matching, so as to realize the matching between the output power of the pump and the sum of the load power of each branch, and effectively improve the system efficiency.

Therefore, the development of controllable variable cross-section hydraulic cylinder is of great significance for improving the load capacity of mobile robots.

Utility model content

The purpose of this utility model is to provide a hydraulic drive system that can effectively improve the efficiency of the mobile robot in view of the deficiencies of the prior art, and can be connected to the high-pressure oil circuit or to the low-pressure oil circuit by selecting different cavities of the multi-stage hydraulic cylinder. To achieve the matching of oil source pressure and load pressure, and finally achieve the purpose of improving the efficiency of the hydraulic system, the variable cross-section hydraulic cylinder and its hydraulic control system.

The utility model realizes above-mentioned object through following technical scheme.

A variable cross-section hydraulic cylinder, including a cylinder barrel, a primary piston and a secondary piston, the primary piston includes a primary piston rod and a cylinder barrel, and the primary piston rod is sealed and installed on the In the cylinder, the secondary piston includes the primary piston rod and the secondary piston rod. The interior of the primary piston rod is hollow, and the primary piston rod serves as the cylinder of the secondary piston rod. The first-stage piston rod is sealed and installed in the first-stage piston rod through the piston ring and the end cover. The first-stage rodless chamber oil passage communicates the external oil passage with the second-stage piston rodless chamber, and the second-stage rod-mounted chamber oil passage communicates the external oil passage with the second-stage piston rod chamber.

A hydraulic control system for variable cross-section hydraulic cylinders, the primary piston rodless cavity, the primary piston rod cavity, the secondary piston rodless cavity, and the secondary piston rod cavity are respectively connected to the oil port C1 of the first switch valve, the second The oil port C2 of the second switch valve, the oil port C3 of the third switch valve and the oil port C4 of the fourth switch valve are connected, the first switch valve and the second switch valve are connected with the first servo valve, and the third switch valve The valve and the fourth switch valve are connected with the second servo valve, the oil port A1 of the first switch valve, the oil port B2 of the second switch valve are connected with the oil port VA1 of the first servo valve, the first The oil port B1 of the first switch valve, the oil port A2 of the second switch valve communicate with the oil port VB1 of the first servo valve, the oil port A3 of the third switch valve, the oil port A3 of the fourth switch valve Port B4 communicates with oil port VA2 of the second servo valve, oil port B3 of the third on-off valve, oil port A4 of the fourth on-off valve communicates with oil port VB2 of the second servo valve, The high-pressure oil inlet P1 of the first servo valve and the high-pressure oil inlet P2 of the second servo valve communicate with the high-pressure oil outlet of the power source composed of the constant pressure variable pump 8 and the safety relief valve 9. The low pressure oil return port T1 of the first servo valve and the low pressure oil return port T2 of the second servo valve communicate with the oil tank.

The switch valve is a two-position three-way electromagnetic switch valve, and the servo valve is a three-position four-way servo valve.

Due to the adoption of the above scheme, the utility model is a linear hydraulic cylinder with a certain load matching ability and a controllable effective area change, which can effectively improve the efficiency of a single pump source-multi-actuator hydraulic drive system with a large change in the load of each actuator. When performing variable load conditions, the maximum output force of the hydraulic cylinder can be achieved by selecting the conduction status between each cavity of the two-stage telescopic hydraulic cylinder and the high-pressure oil circuit and low-pressure oil circuit through the control combination of different electromagnetic switch valves and servo valves. The matching of the load finally achieves the purpose of improving the efficiency of the hydraulic system. The utility model can be used on various small and medium-sized mobile platforms with large variations in the load of each actuator with energy autonomy, such as biped robots, quadruped robots, small unmanned excavators, exoskeleton equipment, etc., and can effectively improve the performance of such equipment. Improve the efficiency of the hydraulic drive system, thereby improving its load capacity, promoting its further practical application, and at the same time realizing energy saving and environmental protection, which has good economic value. In addition, because the cylinder body of the secondary piston is a primary piston, the variable cross-section hydraulic cylinder has a large stroke and a small basic length, which can effectively reduce the installation space of the hydraulic cylinder. The load pressure P _Ln obtained by this selection method will be as close as possible to the oil source pressure P _S , thereby reducing the throttling loss.

Description of drawings

Fig. 1 (a) is the perspective view of the present utility model;

Fig. 1 (b) is a first-stage hydraulic cylinder internal structure diagram of the utility model;

Fig. 1 (c) is the internal structure diagram of the secondary hydraulic cylinder of the present utility model;

Fig. 2 structural principle and hydraulic control system schematic diagram of the utility model;

Fig. 3 is a schematic diagram of the hydraulic cylinder output force of the present utility model;

Fig. 4 is a schematic diagram of matching the output force of the hydraulic cylinder and the load of the utility model;

Fig. 5 is a schematic diagram of the efficiency improvement principle of the utility model applied in a single pump source-multi-actuator hydraulic system.

Detailed ways

Below in conjunction with accompanying drawing, the specific implementation manner of this patent is described in further detail.

As shown in Figure 1, the controllable variable cross-section hydraulic cylinder consists of an end ring 1, a two-position three-way plug-in switch valve 2, a three-position four-way servo valve 3, a cylinder barrel 4, a first-stage piston rod 5, and a second-stage piston rod 6. Composition of end ring 7 with oil holes. Among them, the on-off valve 2 and the servo valve 3 are directly installed on the cylinder 4, corresponding to the first on-off valve SW ₁ , the second on-off valve SW ₂ and the first servo valve SV ₁ in Fig. 2 , and the oil circuit relationship between them It is realized by the integrated oil block 42 on the cylinder 4 as shown in FIG. 1( b ). The oil inlet P1 port and the oil return port T1 port of the first servo valve SV ₁ in Fig. 2 are two external oil circuit interfaces 41 as shown in Fig. 1(b), and the first switch valve SW ₁ and the primary piston The connection of the rodless cavity is realized by the internal oil circuit in the integrated oil circuit block 42, and the connection between the second switching valve SW ₂ and the rod cavity of the first-stage piston is realized by the connecting pipeline 44 in Fig. 1(b), and Fig. 1(b ), the manifold block 42 is provided with a process plug 43 in the oil circuit. The left side of the cylinder 4 is connected with the end ring 1 through threads, and the right side of the cylinder 4 is sealed with the primary piston rod 54 by the end cover 46 , the static sealing ring 45 and the dynamic sealing ring 47 . The primary piston 51 is sealed with the inner wall of the cylinder 4 through the guide ring 52 and the dynamic sealing ring 53 . The primary piston rod 5 is the cylinder barrel of the secondary piston rod 6 , and the right end of the primary piston rod 5 is sealed with the secondary piston rod 6 through an end cover 56 , a static sealing ring 55 and a dynamic sealing ring 57 . The secondary piston 61 realizes sliding and dynamic sealing with the primary piston through a slip ring 62 and a sealing ring 63 . The cavity at the left end of the secondary piston 61 is connected to the oil circuit of the external power system through the oil hole 611, the oil hole 71 on the end ring 7 and the dynamic sealing ring 72, and the cavity at the right end of the secondary piston 61 is connected through the oil hole 612. The oil hole 71 and the dynamic sealing ring 72 on the end ring 7 realize the conduction with the oil circuit of the external power system. In Fig. 2, the cavities connected to the left and right ends of the secondary piston 61 are installed on the frame connected to the end ring 7. The third switch valve SW ₃ , the fourth switch valve SW ₄ and the servo valve SV ₂ are driven by the actuator. Therefore, it is not in the external overall structure diagram shown in Figure 1(a).

The working principle of the utility model: as shown in Figure 2, the variable section of the utility model is realized by four two-position three-way plug-in switch valves 2 and two three-position four-way servo valves 3, high-precision force and Displacement control is realized through two three-position four-way servo valves 3 . The C1 port of the first switching valve _SW1 communicates with the first-stage piston rodless chamber of the variable-section hydraulic cylinder, and the C2 port of the second switching valve _SW2 communicates with the first-stage piston of the variable-section hydraulic cylinder. The rod cavity is connected, the oil port A1 of the first switch valve SW ₁ is connected with the oil port B2 of the second switch valve SW ₂ and the VA1 port of the first servo valve SV ₁ , and the second switch valve SW The oil port B2 of ₂ communicates with the oil port A2 of the second switching valve SW ₂ and the VB1 port of the first servo valve SV ₁ . The port C3 of the third switching valve SW ₃ is in communication with the rod cavity of the secondary piston of the variable cross-section hydraulic cylinder, and the C4 port of the fourth switching valve SW ₄ is in communication with the secondary piston of the variable cross-section hydraulic cylinder. The rod cavity is connected, the oil port A3 of the third switch valve SW ₃ is connected with the oil port B4 of the fourth switch valve SW ₄ and the VA1 port of the second servo valve SV ₂ , and the third switch valve SW The oil port B3 of ₃ communicates with the oil port A4 of the fourth switching valve SW ₄ and the port VB2 of the second servo valve SV ₂ . The P1 port of the first servo valve _SV1 and the P2 port of the second servo valve _SV2 communicate with the high-pressure oil outlet of the power source composed of a constant pressure variable pump 8 and a safety relief valve 9. The T1 port of one servo valve _SV1 and the T2 port of the second servo valve _SV2 communicate with the oil tank 10 .

A load force sensor is set on the secondary piston rod, the first piston forms two surfaces respectively _A1 surface and _Ar surface, and the second piston forms two surfaces respectively _B1 surface and _Br surface, so that in the piston When it stretches out, it has four effective area A _l , B _l , A _l -A _r and B _l -B _r , and when the piston shrinks inward, it has two effective area B _r and A _r , according to the load sensor The measured real-time load is controlled by the electromagnetic switch valve group and the servo valve group, and an appropriate effective area is selected to match the maximum output force of the hydraulic cylinder with the load force. The specific control method is as follows:

Let the effective area be Ae, let the control amount of the on-off valve be x _k , (k=1, 2, 3, 4), the above x ₁ , x ₂ , x ₃ , x ₄ represent the first on-off valve, the second On-off valve, the third on-off valve and the fourth on-off valve, x _k = 1 means that the electromagnetic on-off valve is in the left position, x _k = 0 means that the electromagnetic on-off valve is in the right position, and the control amount of the servo valve is u _k (k = 1, 2), u ₁ and u ₂ points represent the first servo valve and the second servo valve, u _k =-1 means the servo valve is in the left position with the maximum opening, u _k =0 means the servo valve is in the middle position, u _k = 1 means that the servo valve is in the right position with the largest opening;

When the electromagnetic switch valve and the servo valve are in different control states, corresponding to different effective area, the effective area control is shown in the table below;

When the output pressure of the constant pressure variable pump 8 is constant, the output force of the hydraulic cylinder can be approximated by for:

F _O = P _s A _e

Among them, P _s is the outlet pressure of the constant pressure variable pump, and F _O is the output force of the hydraulic cylinder.

By selecting different control states of the three-position four-way servo valve 3 and the two-position three-way plug-in switch valve 2, different effective area A _e can be obtained, thereby obtaining different hydraulic cylinder output forces. By designing the action area, the output force of hydraulic cylinders with different distributions can be obtained, such as the output force of hydraulic cylinders with uniform distribution, as shown in Figure 3. For different loads, different hydraulic cylinder output forces are selected to match them. The schematic diagram of load matching is shown in Figure 4.

The principle of using the variable cross-section hydraulic cylinder with controllable effective area to improve the efficiency of the drive system is as follows. The basic principle block diagram of the single-pump source and multiple-actuator hydraulic drive system of the mobile robot is shown in Figure 5. Assuming that the variable load of the nth valve-controlled cylinder drive branch is F _Ln and the load pressure is P _Ln , then:

P _Ln ＝F _Ln /A _e

P _Ln ＝F _Ln /A _e When a constant pressure oil source is used (this type of oil source is usually used on mobile robots), the output pressure of the constant pressure oil source is set to P _S , and different effective areas correspond to different maximum outputs of hydraulic cylinders Force P _S A _e (that is, the output force of the hydraulic cylinder when the servo valve opening is the largest, to simplify the description, the throttling pressure drop of the valve when the servo valve opening is the largest is omitted here, which does not affect the analysis results essentially). From the above, 7 effective action areas A _e can obtain 7 different maximum output forces of hydraulic cylinders, which is set as F _k (k=1~7). Through reasonable size design, the maximum output force F _k of hydraulic cylinders with different distributions (k=1-7) can be obtained, as shown in Figure 3, the maximum output force of hydraulic cylinders with uniform distribution.

Load matching and effective area selection method: the load force sensor obtains the actual load F _Ln in real time, compares it with the maximum output force of the 7 hydraulic cylinders, and determines that the actual load falls between the two maximum output forces, such as F _k-1 < F _Ln < F _k . If the hydraulic cylinder extends outward, the maximum output force should be greater than the actual load to drive the load to extend outward, then choose F _k as the maximum output force, and the effective area corresponding to F _k is the effective area to be selected; if you choose F _k As the maximum output force, by detecting the tracking error of the hydraulic cylinder at the next two moments, if the tracking error (tracking error: the difference between the expected displacement and the actual displacement of the piston rod of the hydraulic cylinder) further expands, it means that the selected The maximum output force is not large enough to meet the dynamic tracking requirements, then further increase the maximum output force of the hydraulic cylinder, select F _k+1 as the maximum output force, and then repeat the above steps until the tracking error no longer expands, and at this time select the maximum The effective area corresponding to the output force is the effective area to be selected. Similarly, if the hydraulic cylinder retracts, the maximum output force should be smaller than the actual load to pull the load back, then select F _k-1 as the maximum output force, and then repeat the above steps until the tracking error no longer expands, at this time The effective area corresponding to the selected maximum output force is the effective area to be selected; if F _k-1 is selected as the maximum output force, even if the opening of the servo valve is the largest at the next two moments, the system tracking error will further expand, It means that the selected maximum output force is not small enough to meet the dynamic tracking requirements, then further reduce the maximum output force of the hydraulic cylinder, select F _k-2 as the maximum output force, and the effective area corresponding to F _k-2 is the one to be selected effective area. The load pressure P _Ln obtained by this selection method will be as close as possible to the oil source pressure P _S , thereby reducing the throttling loss. This method can meet the requirements of dynamic tracking accuracy while achieving a high degree of load matching, that is, it can not only realize the system function requirements, but also improve the driving efficiency.

Let's analyze the power throttling loss of the branch:

The power throttling loss of the nth branch is:

ΔW _n =P _s ·Q _n -P _Ln ·Q _n

＝(P _s -P _Ln )·Q _n

=ΔP _n Q _n

When the load F _Ln changes, by adjusting A _e , P _Ln can be made close to P _s , that is, ΔP _n can be less than a relatively small pressure drop ΔP _min , namely:

ΔW _n <ΔP _min Q _n

The total power throttling loss of the drive system is:

$\mathrm{<span\; class="notranslate">\; \&Delta;W</span>}<span\; class="notranslate">\; =</span>{\mathrm{<span\; class="notranslate">\; \&Sigma;</span>}}_{\mathrm{<span\; class="notranslate">\; 1</span>}}^{\mathrm{<span\; class="notranslate">\; N</span>}}\mathrm{<span\; class="notranslate">\; \&Delta;</span>}{\mathrm{<span\; class="notranslate">\; W</span>}}_{\mathrm{<span\; class="notranslate">\; no</span>}}<span\; class="notranslate">\; <</span>\mathrm{<span\; class="notranslate">\; \&Delta;</span>}{\mathrm{<span\; class="notranslate">\; P</span>}}_{\mathrm{<span\; class="notranslate">\; min</span>}}{\mathrm{<span\; class="notranslate">\; \&Sigma;</span>}}_{\mathrm{<span\; class="notranslate">\; 1</span>}}^{\mathrm{<span\; class="notranslate">\; N</span>}}{\mathrm{<span\; class="notranslate">\; Q</span>}}_{\mathrm{<span\; class="notranslate">\; no</span>}}<span\; class="notranslate">\; =</span>\mathrm{<span\; class="notranslate">\; \&Delta;</span>}{\mathrm{<span\; class="notranslate">\; P</span>}}_{\mathrm{<span\; class="notranslate">\; min</span>}}{\mathrm{<span\; class="notranslate">\; Q</span>}}_{\mathrm{<span\; class="notranslate">\; the\; s</span>}}$

Q _s can be adjusted in real time through the variable adaptive structure of the constant pressure variable pump, that is, the pump provides as much flow as the system needs at each moment. Therefore, it can be seen from the above formula that the power throttling loss of the drive system mainly depends on ΔP _min , that is, the closer the load pressure P _Ln of each branch is to the outlet pressure P _s of the pump, the smaller the power loss of the system.

From the above analysis, it can be seen that compared with the traditional hydraulic cylinder with fixed cross-section and the multi-stage telescopic hydraulic cylinder _that can only change the cross-section at a fixed position, as the load changes, the P _Ln of each branch changes greatly. greater than the maximum load pressure P _Lmax that occurs during the operation of the entire system, resulting in a large change in the throttling pressure drop ΔP _n of each branch and a large difference between them, it cannot be unified to a smaller pressure drop like the utility model in the design. The pressure drop ΔP _min , so its power throttling loss is much larger than the power throttling loss ΔP _min Q _s of the utility model; and the application of this patent can adjust the effective area A _e in real time according to the change of the load F _Ln , so that the load pressure P _Ln is close to or equal to the pump source pressure, reducing power loss, thereby improving system efficiency.

Another feature of the utility model is: because the cylinder body of the secondary piston is a primary piston, the controllable variable cross-section hydraulic cylinder has a short basic length and a long stroke, and the required installation space is small.

Claims (3)

1. a controlled variable cross section oil hydraulic cylinder, it is characterized in that: comprise cylinder barrel, first stage piston and second piston, described first stage piston comprises first stage piston bar and cylinder barrel, described first stage piston bar is arranged in described cylinder barrel by piston ring and end bracket enclosed, described second piston comprises described first stage piston bar and second piston rod, the boring of described first stage piston bar, described first stage piston bar is as the cylinder barrel of second piston rod, described second piston rod is arranged in described first stage piston bar by piston ring and end bracket enclosed, described second piston rod inside is provided with described secondary rodless cavity oil circuit and secondary rod chamber oil circuit, outside oil circuit is communicated with described second piston rodless cavity by described secondary rodless cavity oil circuit, outside oil circuit is communicated with described second piston rod chamber by described secondary rod chamber oil circuit.

2. the hydraulic control system of a controlled variable cross section oil hydraulic cylinder, for controlling variable cross section oil hydraulic cylinder as claimed in claim 1, it is characterized in that, first stage piston rodless cavity, first stage piston rod chamber, second piston rodless cavity and second piston rod chamber respectively with the hydraulic fluid port C1 of the first switch valve, the hydraulic fluid port C2 of second switch valve, the hydraulic fluid port C3 of the 3rd switch valve is connected with the hydraulic fluid port C4 of the 4th switch valve, described first switch valve is connected with the first servovalve with second switch valve, described 3rd switch valve is connected with the second servovalve with the 4th switch valve, the hydraulic fluid port A1 of described first switch valve, the hydraulic fluid port B2 of described second switch valve is communicated with the hydraulic fluid port VA1 of described first servovalve, the hydraulic fluid port B1 of described first switch valve, the hydraulic fluid port A2 of described second switch valve is communicated with the hydraulic fluid port VB1 of described first servovalve, the hydraulic fluid port A3 of described 3rd switch valve, the hydraulic fluid port B4 of described 4th switch valve is communicated with the hydraulic fluid port VA2 mouth of described second servovalve, the hydraulic fluid port B3 of described 3rd switch valve, the hydraulic fluid port A4 of described 4th switch valve is communicated with the hydraulic fluid port VB2 of described second servovalve, the high pressure oil inlet P 1 of described first servovalve is communicated with the high-pressure oil outlet of the power source that the high pressure oil inlet P 2 of described second servovalve is formed with safety overflow valve (9) with constant pressure variable displacement pump (8), the low pressure oil return inlet T 1 of described first servovalve is communicated with fuel tank with the low pressure oil return inlet T 2 of described second servovalve.

3. the hydraulic control system of controlled variable cross section oil hydraulic cylinder according to claim 2, it is characterized in that: described switch valve is two-bit triplet electromagnetic switch valve, described servovalve is 3-position 4-way servovalve.