ELECTRO-HYDRAULIC ACTUATOR

The present invention relates to electromagnetic actuators and hydraulics, and more particularly relates to an apparatus for providing a combined electro-hydraulic actuator. Electromagnetic actuators are typically used as linear motors to provide a large force on an armature which is positioned within a channel formed by the motor. The use of electromagnetic actuators is desirable when a large force is required to be exerted in a linear direction. Hydraulic actuators perform a similar function to electromagnetic actuators in that they are capable of controlling mechanical motion. However, the latter uses high pressure hydraulic pumping which requires frequent maintenance and is energy inefficient. There are certain advantages in using an hydraulic actuator such as the capability for the pressure of the actuator to be controlled in order to lock it in position.

Accordingly, the present invention aims to combine the two separate actuators in order to provide the advantages of both actuators in a single system.

This is achieved by providing an electro-hydraulic system comprising a electromagnetic actuator and a hydraulic actuator relatively arranged such that the movement of the hydraulic actuator is controlled by the movement of electromagnetic actuator.

Preferably, a pump arrangement is provided which serves as an interface between the electromagnetic actuator and the hydraulic actuator and is configured to enable both forward and reverse movement of the hydraulic actuator by way of electrically controlled valves.

The electrohydraulic system can be incorporated into an integral self-contained unit and a hydraulic actuator is provided with its own electrically controlled pump. Advantageously, this obviates the need for hydraulic piping and auxiliary hydraulic devices to be connected to the system and enables the system to be controlled electrical/electronic inputs only. hi order that the present invention be more readily understood, embodiments thereof will be described with reference to the accompanying drawings in which:

Fig 1 is a schematic diagram showing the basic principle of operation according to a preferred, embodiment of the present invention; Fig. IA is a schematic diagram showing an alternative arrangement of the valves of Fig.

1.

Fig 2 is a detailed diagram of one state of a valve used in Fig 1 ;

Fig 3 is a detailed diagram of another state of the valve used in Fig 2;

Fig 4 is a detailed diagram showing an upper part of the pump in Fig 1 ; and Fig 5 is an overall diagram showing the integrated design of the apparatus in Fig 1. Fig. 6 is a modification to the embodiment shown in Fig 1. Fig. 7 is a timing diagram showing the position of various valves shown in Fig. 6. As shown in Fig 1, a linear electromagnetic actuator 1 operates a reciprocating pump 2.

The type of linear actuator most suited to this embodiment comprises a non-magnetic shaft Ia arranged to move along a linear actuator cylinder Ib. The shaft Ia is attached to an armature Ic which moves within the cylinder Ib as a result of the magneto motive force created by the switching of current in a plurality of current carry conductor coils Id which are arranged along the periphery of the cylinder Ib. Between each coil Id is an air gap Ie and this configuration is repeated along the periphery. The switching of the current in the coils can be digitally controlled by for example a computer and therefore it is possible to control the movement of the shaft Ia along the cylinder Ib. It will be appreciated that alternative forms of linear actuator may be utilised as long as the shaft can be controlled in the manner described above. The distal end of the shaft Ia away from the armature Ic is arranged to move through an opening in the pump 2 and this end of the shaft Ia is attached to a pump cylinder 2a.

The cylinder 2a of the pump 2 has an electrically controlled valve arrangement 3 comprising two non-return valves 3 a, along with two shut-off valves 3b in series as shown in Fig 1. The non-return valves 3 a admit flow in opposite directions. In operation only one of the shut-off valves 3b is open and the other closed. This makes the pump deliver fluid in opposite directions. Where the size of the pump piston is not large enough to accommodate these integral non-return and shut-off valves, they may be placed outside independently as shown in Fig Ia. In this case, standard solenoid or other electrically operated shut-off valves 30b are used.

Also provided are two diaphragm based small sized accumulators 4 which are incorporated in the integral design, to allow for slight volume changes due to compressible nature of real fluids and also account for temperature expansion/contraction of the fluid.

The delivered fluid from the pump directly enters a double acting hydraulic reciprocating actuator 5. This will cause hydraulic piston 6 to move by an amount depending on the amount of fluid delivered and the relative cross-sectional areas of the pump 2 and the hydraulic actuator 5. The force supplied by the hydraulic actuator 5 is amplified by the area ratio of the pump 2 and hydraulic actuator piston 6. The relationships are derived under theoretical analysis. In order that the continuity is satisfied for incompressible fluid, the dimensions of the piston rod 6 and cylinder bore 7 are to be chosen appropriately. Furthermore this continuity can extend for

compressible fluids by computing the average values and causing the actuator 5 to respond to maintain the continuity by matching the requirements by way of employing a fast response (eg 1 ms) actuator.

The switching on and off of the shut-off valves 3b enables reverse/forward action of the hydraulic actuator 5. As shown in more detail in Fig 2, 3 and 4, this is achieved through use of electrically controlled protruding members 10 which allow the valve body to slide to open/close positions inside the pump piston cylinder 2a. In particular, with reference to Figs 2 and 3, the non-return and shut-off valve is incorporated into the cylinder 2a. A spring 60 retains ball 70 of the non-return valve 3 a and the shut-off valve 3b is shown in the open state in Fig 2. Numeral 9 represents the pressure side of the cylinder 2a. Numeral 8 represents the other side of the cylinder 2a and fluid enters the valve as shown from this side. The condition when the shut-off valve 3b is in a shut-off state is shown in Fig 3.

It will be appreciated that as well as being designed as separate components, the system may be designed as an integral self-contained unit as shown in Fig 5. The advantage of this type of integral unit is that it requires only electric wiring input into the unit so as to control the electromagnetic actuator and pump and avoids the use of hydraulic piping. That is, the hydraulic actuator forming part of the integral unit is provided with its own electrically controlled pump obviating the need of fluid reservoirs motors and other pumps which are required as auxiliary units for conventional hydraulic systems. Therefore, high pressure hydraulic piping can be dispensed with and pressure drops which are usually encountered in pipelines can be avoided. It will be appreciated that it is possible to adapt the system to enable more than one hydraulic actuator to be controlled by the same pump.

In addition this type of system will produce very little noise compared to a conventional system which uses hydraulic auxiliary units. The use of the linear electromagnetic actuator in combination with the hydraulic combination utilises the benefit of the high forces generated by the linear actuator and that of the hydraulic actuator, namely loclcable in any position to result in an efficient and easily controllable electro-hydraulic system.

Additionally, the hydraulic amplification for the force generated by the electric actuator makes the application of the linear electric actuator more efficient for a given force and makes it possible to achieve greater application force for any given design of the electric actuator.

In a modification to the embodiment shown in Fig. IA a valve arrangement 300 is provided which results in a higher overall efficiency of the electro-hydraulic system. The modification is shown in Fig. 6. Features which are the same as those used in Fig. IA will be referred to with like numerals. The electromagnetic actuator 1 as described in Fig. IA is provided and a shaft 100a protrudes from one side of the cylinder Ib of the actuator 1. The shaft 100a differs to that in Fig IA in that the distal end is not directly attached to a pump cylinder. Instead the distal end of the shaft 100a in mounted in order to control the movement of a mechanical framework comprising three members 110, 111, 112. The distal end of shaft 100a is attached to first member 110 at one end thereof. The other end of the first member 110 is attached to one end of the second member

111, and the other end of the second member 111 is attached to one end of the third member

112. The framework is arranged such that movement of the shaft 100a in a first direction will cause movement of the second member 111 in the first direction.

A first compression spring 114 is positioned to contact the third member 112 when the shaft 100a comes the end of its stroke and is fully extended. A second compression spring 115 is positioned to contact the first member 110 when the shaft 100a approaches the end of its stroke within the cylinder Ib. The springs 114,115 may be helical or equivalent types. The kinetic energy of the armature Ic of the linear actuator 1 driving a pump 500 at the extremities of the stroke is stored and released. A middle portion of the second member 111 forms part of a double acting pump 200 and the member 111 is arranged to move within a cylinder 200a forming the housing of the pump 200. An actuator bore 220b is attached to the second member 111 and the double acting pump 200 includes a number of ports such that the volume of fluid moved within the cylinder 200a as a result of the movement of the second member 111 and thus the bore 220b, fluid is able to exit the respective ports.

A valve arrangement 300 is arranged to operate on the basis of the double acting pump 200. Valve 301 and 302 are delivery and suction valves with two possible positions. Valve 303, 304, 305 are delivery and suction valves with three possible positions.

The pump 500 comprises a piston rod 600 and a cylinder bore 700 similar to that shown in Fig. 1 and IA. The valve 305 is arranged to compensate low pressure volume which is described below.

Fig. 7 shows the valve position timing for valves 301,302,303, and 305 over one pump cycle. The bold line indicates the particular position of each respective valve. With reference to

Fig. 6 and 7, displacement x=0 when the double acting actuator bore 220b is in a middle position within the cylinder 200a. Displacement X=L when the electromagnetic actuator shaft 100a has fully extended and x=-L when the shaft 100a is at the other extremity.

The compression springs 114,115 come into action at the ends of the pump strokes. The KE of the pump piston and armature is stored in the spring during the overshoot of pump piston 600 and returned to the masses for the reverse stroke.

During the pump piston overshoot at the end of strokes, delivery and suction connections to and from pump piston are shut off by mid-position (position 2) of valve 303. During the pump piston overshoot, both sides of cylinder across the pump piston are short circuited using position 2 of Valves 301 and 302.

Positions 1 and 3 of Valve 303 are set to give one directional flow for the double acting pump 200.

Positions 1 and 3 of Valve 304 provide forward and reverse motion of load piston. The middle position of Valve 304 locks and load piston. Compensation low pressure volume is used to accommodate the volume difference of the piston causes different swept volumes on either side of the actuator piston. The volume capacity of this should be about 3 to 4 times the volume of the piston rod in the main actuating cylinder. This reservoir will also allow for temperature and compressibility effects. Although this volume is shown integral with the actuator cylinder, it may be placed away where that is required. Safety pressure valves are not shown in Fig. 6.

With the arrangement in Fig. 6, the kinetic energy of the armature 1 C of the actuator 1 driving the pump 200 at the extremities of its stroke is stored and released while the delivery and suction valves 303, 304 are short circuited. This action causes the electromagnetic actuator 1 to generate an output force when it has higher velocity. The effect of this results in a highly efficient actuator.

Furthermore with the modification the following advantages can be provided.

- The use of return springs results in a high efficiency actuator.

- It is possible to use external standard control valves to achieve the electro-hydraulic actuation without the need for in-built valves shown in Fig. 1. - The use of compensating volume, either integrated with the actuator as shown in Fig.

6 or located separately makes use of actuators which have a piston rod protruding only on one side.

It is feasible to integrate the components in mono-block design avoiding the use of external fluid or piping.

- The hydraulic actuation is utilised as an instantaneous hydraulic lever unlike conventional systems where the energy is converted and stored as high pressure fluid and released in a separate step.

- The electro-hydraulic actuator has a high static resistance which follows from the use of a hydraulic load cylinder, when its inlet and outlet valves are shut off and the fluid within the cylinder trapped.