WO2001081761A1

WO2001081761A1 - A coupling and a method for equalizing variations in the volume flow in a hydraulic engine

Info

Publication number: WO2001081761A1
Application number: PCT/FI2001/000398
Authority: WO
Inventors: Wolfgang BACKÉ; Matti Vilenius; Janne UUSI-HEIKKILÄ
Original assignee: Backe Wolfgang; Matti Vilenius; Uusi Heikkilae Janne
Priority date: 2000-04-27
Filing date: 2001-04-25
Publication date: 2001-11-01
Also published as: FI110960B; FI20000987A; FI20000987A0; AU2001258434A1

Abstract

The invention relates to a coupling for equalizing variations in the volume flow of a hydraulic engine, the hydraulic engine (M) comprising at least one periodically operated displacement chamber (4) and being arranged to suck pressurized medium into this displacement chamber (4) from the suction side (T) of the hydraulic engine, to compress the pressurized medium sucked into the displacement chamber (4) to a working pressure and to supply the compressed pressurized medium out from the displacement chamber (4) to the pressure side (P) of the hydraulic engine. For damping of variation caused by the compression, the coupling is arranged to supply pressurized medium into the chamber (4), to compress the pressurized medium sucked into this displacement chamber (4).

Description

A coupling and a method for equalizing variations in the volume flow in a hydraulic engine

The invention relates to a coupling for equalizing variations in the volume flow in a hydraulic engine according to the preamble of the appended claim 1. The invention also relates to a method for equalizing variations in the volume flow in a hydraulic engine according to the preamble of the appended claim 9.

Hydraulic engines operating by the principle of displacement have a large power density, because they are suitable for use under very high pressures, typically even 300 to 400 bar and more. In a hydraulic pump, mechanical energy, such as driving torque and rotational frequency, is converted to hydraulic energy, i.e. pressure and volume flow of a pressurized medium. In a hydraulic motor, hydraulic energy is converted to mechanical energy again. The rotational frequencies typically range from 500 to 10,000 rpm. The torque given by the hydraulic motor depends on the displacement volume in the motor and the pressure difference effective over it. Further, the output depends on the volume flow and the pressure difference across the motor. Slowly rotating engines, whose rotational speeds typically range from 0.5 to 1 ,000 rpm, are designed in such a way that they give a large torque even at low rotational speeds. These engines are called LSHT engines (Low Speed High Torque). Several hydraulic engines can be used both as pumps and as motors, thanks to their principle of operation. Pumps operating by the principle of displacement comprise mechanically sealed cham- bers used as displacement volumes, into which the fluid is sucked, for example by means of a motion of a piston, from the suction side or from the inlet side, and by means of which the fluid is transferred to be supplied to the pressure side or the outlet side, for example by means of a motion of the piston. In motors, the operation is reverse.

Large drops or increases in the pressure in the chambers produce strokes in the body of the engine and thereby body sounds which can be heard as disturbing noise emitted to the environment. Furthermore, noise is caused by variation in the volume flow, which is due to uneven distribution of said volumes in relation to the angle of rotation of the engine. These volumes are provided, for example, in an axial or radial piston pump or in a corresponding motor. Variations in pressure also lead to resonance in other structures related to the system, which fur- ther increases the noise and is often also the most significant source of the noise.

German application publication 1703210 presents a solution for reducing variations in the volume flow produced by a hydraulic engine, wherein a very small or even number of pistons can be used. As an example, a two-piston axial piston pump has been used. On the other hand, variations in the volume flow can be reduced by increasing the number of pistons, wherein typically 7 to 11 pistons are used; however, this will result in a structure which is larger, more complex and more expensive. Typically, also an odd number of pistons is used to achieve a volume flow which is as steady as when using a double but even number of pistons.

The solution presented in the application publication 1703210 comprises a pair of auxiliary pistons which operate in a synchronized manner in parallel with the working pistons and supply the volume flow to the outlet side, when the volume flow supplied jointly by the working cylinders is below a determined average. The radial auxiliary pistons receive the volume flow from the outlet side when the volume flow jointly supplied by the working cylinders exceeds a determined average. Variations in the volume flow are dependent on the angle of rota- tion of the driving shaft, wherein the auxiliary pistons are also coupled onto the driving shaft for synchronization. The displacement volume required by the auxiliary pistons is considerably smaller when compared with the working cylinders. Thus, it is easy to place them in the engine; for example, a four-piston engine will only require one auxiliary piston. For the sake of power balance, however, two opposite auxiliary pistons are preferred.

In the alternative when a hydraulic engine is used as a motor, the volume of the auxiliary piston is coupled to the pressurized inlet side, for example by means of controllable valve means, such as a 3/2 directional control valve. In motor use, the auxiliary pistons can be used to equalize variations in the output torque. Driven by the pressurized medium, the auxiliary piston gives the required auxiliary torque at a desired moment during the motion of the working cylinder. To remove the pressurized medium, however, the auxiliary piston applies the torque force, wherein the auxiliary piston will equalize variations in the torque. The friction varies in different directions of motion of the auxiliary piston, but it will cause a relatively small variation in the torque when compared to a situation in which no auxiliary pistons are used for the compensation described above.

The problem is that fluctuations in the pressure and the volume flow also take place on the suction side of the hydraulic engine used as a pump, depending on the angle of rotation. Moreover, a pulsation remains in the volume flow on the pressurized output side, causing compression of the pressurized medium particularly in the phase when unpressurized fluid sucked into the displacement chamber is started to be compressed against a load, such as a pressure valve, wherein its pressure is simultaneously increased. The pressure is increased to the working pressure to be used, which can be adjusted with a pressure controller in connection with the hydraulic engine or with a pressure control valve in connection with the system. In the valve, the pressure is typically effected on a movable, spring-actuated seat or gate, wherein the maximum pressure can be changed by adjusting the tension of the spring. In hydrostatics, as known, an external force generates in a medium a pressure which is spread in all directions in the fluid. The magnitude of the pressure is obtained by dividing the force by the surface area under the pressure.

However, entrained air in the fluid causes compression of the fluid when the pressure increases, wherein the volume of the fluid is reduced. Thus, quick pressure fluctuations cause pressure strokes caused by changes in the volume. The compression is represented by a constant depending on the pressurized medium, being, for example, for mineral oil about 0.7 to 0.8 % and for water about 0.45 % per 100 bar. In a hydraulic pump, the fluid must be compressed before raising the pressure as well as before the sufficient working pressure is achieved and some of the volume flow is shifted to the pressure side. The greater the volume to be compressed, the longer it takes to achieve the working pressure, and the greater the interference.

It is an aim of the present invention to eliminate the above-mentioned problems. The aim is to present a coupling, whereby it is possible to compensate for the delay caused by the compression. The coupling according to the invention is characterized in what will be presented in the characterizing part of claim 1. The method according to the invention is characterized in what will be presented in the characterizing part of claim 9.

The invention is based on the idea that a pressurized medium which is substantially under the working pressure is supplied to the displace- ment chamber from a pressure source, such as a pressure accumulator. The feeding takes place just before starting to compress the pressurized medium in the chamber, subject to the pressure of the suction side, for example by means of a working piston. By supplying a volume flow from the pressure accumulator, the medium can be quickly brought to the working pressure, wherein not only the working piston is used for the compression, and the medium, under the working pressure, can be supplied earlier from the displacement volume. Thus, the fluctuations become considerably smaller than in prior art.

According to an advantageous embodiment of the invention, the pres- sure accumulator is loaded at the working pressure by choking from the pressure side of the hydraulic engine, and the supply of the medium into the chamber is controlled by valve means, mechanically or electrically. According to an advantageous embodiment of the invention, the hydraulic engine comprises, for compensation of the volume flow, aux- iliary pistons, whose motion is utilized to close and open the connection between the pressure accumulator and the volume. Thus, the auxiliary pistons are used as said valve means. The operation is based on the fact that the supply of the medium from the pressure accumulator can be synchronized with the motion of the auxiliary piston. The auxiliary piston can also be used as a mere movable gate for the valve means.

According to a preferred embodiment of the invention, auxiliary pistons are also used for reducing fluctuations in the volume flow on the suction side. These fluctuations depend, in a corresponding manner, on the angle of rotation, because the same working pistons are used to suck the medium to be compressed. Thus, fluctuations in the pressure are also reduced, which cause cavitation and wearing of the engine, as well as resonance and noise in the structures, which are due to the fluctua- tions.

To provide compensation, i.e. auxiliary pistons, simultaneously on both the suction side and the pressure side, even a double number of pistons is required. The pair of auxiliary pistons thus operates with another, crosswise pair of auxiliary pistons in the opposite phase. Thus, their manufacture and placement in the engine becomes considerably more complex, resulting in larger and more massive devices. To eliminate the problem, a novel piston structure is presented for compensating fluctuations in the volume flow on the suction and pressure sides. The compensating piston has the considerable advantages of space saving, weight saving and an integrated structure. The principle is to form two displacement chambers in the same auxiliary piston, wherein the piston moves in the cylinder volume, simultaneously forming two separate chambers which operate in opposite phases.

By combining the use of the medium supply and the compensation on the output side, a considerably steady volume flow is achieved. A pulsation remained in the compensated volume flow can be eliminated by compression. By compensation on the inlet and outlet sides, a hydraulic engine with a lower noise level is achieved.

In the following, the invention will be described in more detail by means of an advantageous embodiment with reference to the appended drawings, in which Fig. 1 shows a two-piston pump according to prior art, comprising two compensation pistons,

Figs. 2a to 2d illustrate the operation of the compensation pistons of Fig. 1 to equalize the volume flow,

Fig. 3 shows a two-piston pump according to an advantageous embodiment, comprising two compensation pistons,

Fig. 4 illustrates, in a chart, the variation of a volume flow in a two-piston pump and a pulsation caused by compression,

Fig. 5 shows a two-piston pump according to an advantageous embodiment, comprising a coupling for eliminating the pulsation, and

Fig. 6 shows an alternative advantageous embodiment to the coupling according to Fig. 5.

Figure 1 shows a simplified view on a hydraulic engine M, known as such, which is used as a pump by the principle of displacement and which comprises two auxiliary pistons for the compensation of the volume flow. The pump is a two-piston radial piston pump used by means of a shaft 1. On the shaft 1 , a wobbler 2 is arranged, along which the piston 3 glides. The pump body B is provided with two displacement volumes 4 by means of cylinders 5, a piston 3 reciprocating in the cylinder 5 in a sealed manner. The chamber 4 is connected to the suction side T to suck a fluid pressurized medium by the effect of underpressure into the chamber 4 which is expanded by the movement of the piston 3. During the suction, the suction valve 6 is open and the pressure valve 7 is closed. In the lower dead point of the piston 3, the suction is changed to compression, wherein the pressure of the fluid substantially rises to the working pressure and the fluid is supplied to the pressure side P. During the suction, the suction valve 6 is closed and the pressure valve 7 is open. The maximum of the working pressure and the pressure on the pressure side P can be controlled, for example, by a pressure regulating valve 8, such as a pressure control valve. The working pressure can also be determined by another load coupled to the pressure side P.

In the upper dead point of the piston 3, the supply is changed to suction again. The shaft 1 is rotated, for example, with an electric motor, wherein the pump operates in a periodic, sinusoidal manner. The opposite pistons 3 are arranged to operate in opposite phases, wherein one is feeding the fluid while the other is sucking it. The wobbler 2 of Fig. 1 is arranged in such a way that each piston 3 operates one suction and one feeding during a rotation of the shaft 1.

Figure 1 also shows two displacement chambers 9 of prior art, formed with a cylinder 10 in which a piston 11 reciprocates in a sealed manner. The piston 11 is controlled by means of a cam 12 fitted on the shaft 1, the bearings 13 of the piston 11 gliding along the cam 12. The cam 12 is arranged to move the piston 11 to and fro twice during one rotation of the shaft. The volume displaced by the piston 11 is about 21 % of the volume displaced by the piston 3, to equalize the volume flow Q. In a corresponding manner, in a four-piston pump, the volume is only about 4 %, wherein the pistons 11 perform four strokes during a rotation. The auxiliary chamber 9 is synchronized to receive the volume flow to be supplied from the working chamber 4 to the pressure side P via a channelling 28, when the joint supply of the working chambers 4 is greater than the average or a selected level. The auxiliary chamber 9 is synchronized to discharge a volume flow from the auxiliary chamber 9 back to the pressure side P, when the joint supply of the working chambers 4 is smaller than the average or selected level. By the shape of the profile of the cam 12, it is possible to time the operation mechanically and to control the motion of the piston 11 to equalize the volume flow. Two opposite auxiliary pistons 11 move at the same stroke.

Figure 2a shows, in a chart, the stroke h of two working pistons 3 in more detail as a function of the rotation angle φ of the shaft 1 , wherein φ=2π corresponds to one rotation. Figure 2b shows the variation of the volume flow Q produced by the working pistons 3 together on the pressure side P, as a function of the rotation angle φ, and it is seen that the volume flow is greater than the average Q in the range φι — φ₂ and smaller in the range φ₂ — (p₃ as well as in the range φ₀ — φ-i. The same is repeated at intervals of half a rotation. The auxiliary pistons 11 receive an extra volume flow from the pressure side P in the range q^ — φ₂, indicated with diagonal lines, and then discharge it to the pressure side P in the next ranges φ₂ — ς>₃ and φ₀ — φ^. Figure 2c shows the position of the auxiliary pistons 11 as a function of the rotation angle φ. Figure 2d shows the stroke h of the auxiliary pistons 11 as a function of the rotation angle φ of the shaft 1 , the stroke being produced and synchronized by means of the shaped cam 12.

When the hydraulic engine is used as a motor, the compensation cou- pling on the pressure side T is implemented in the same way as the channels 27a and 27b of Fig. 3, but the channel 27 is coupled to the chamber 9a of Fig. 3 or to the chamber 9 of Fig. 1. It is obvious that control valves, such as pressure control valves and check valves, can be placed in the channels, wherein the operation is not dependent on the suction valve 6 or on the placement of the channels, the valves being electrically controlled. It must be possible to turn off the compensation, particularly when the inlet side T is used as the suction side of the pump.

The coupling on the inlet side T of the pump to equalize the volume flow is presented in a reduced manner in Fig. 3, wherein a two-chamber auxiliary piston 11 is applied. The operation of the auxiliary pistons 11 complies with Figs. 2a to 2d. Alternatively, each auxiliary piston 11 takes care of the compensation of either the outlet side or the inlet side separately, wherein in the following description, only one of the cham- bers 9a, 9b is in use. In the hydraulic engine, at least two separate displacement chambers 9a and 9b are provided for the same stroke.

In the example, a piston 11 equipped with a rod part and a piston part moves in a cylinder 10 which is provided with a chamber in which the rod part moves in a sealed manner. The piston part fixed to the rod part divides the chamber into two displacement chambers, wherein when the chamber 9a is expanded during the suction, the chamber 9b is reduced during the supply, and vice versa. It is also feasible that the piston 11 comprises several piston parts which are fitted to move simultaneously in two or more chambers separated from each other. The suction side T can be compensated with a pair of auxiliary pistons, which, referring to Fig. 1 , can be controlled by the same shaft 1 and the same cam 12 whose cross section is substantially oval. The second auxiliary pistons are thus at an angle of 90° to the direction of rotation, in relation to the auxiliary pistons 11 , wherein the phase is reverse. As the cam 12 controls the auxiliary pistons 11 into four strokes during a rotation, the second auxiliary pistons must be placed at an angle of 45°, wherein the cam 12 is substantially a square whose corners are strongly rounded and whose sides are concave. It is also possible to arrange the crankcase of the hydraulic engine M as the chamber 9b. This gives the advantage that the changes in the volumes of the chamber 9a and the crankcase are equal, wherein the volume flows controlled by the chambers 9a, 9b are also equal.

With reference to Fig. 3, the chamber 9b is connected to the suction side T via the channel 27. One piston 11 and the chamber 9b can be used to control at least two pistons 3 with different phases, wherein the channels 27a and 27b are led to the inlet side before the suction valves 6. Thus, by the pressure effect of the compressing piston 3, one suction valve 6 is closed and the other is open as a result of the suction phase, wherein the volume flow discharged by the chamber 9b is led to the sucking piston 3 in the range φi — φ₂, as shown in Fig. 2b. The suction phase is illustrated by the declining curve 3 in Fig. 2a, and the change in the volume flow Q on the suction side T is illustrated in Fig. 2b. The motion of the piston 11 is shown in Fig. 2c, wherein the piston 3 receives an additional volume flow more evenly according to the curve Q_M. Pressure changes are reduced, and particularly the production of cavitation is avoided, wherein the pressure cannot be reduced to the evaporating pressure. When the sucked volume flow is smaller than the average Q_M in the range φ₀ — φ^, φ₂ — ψs, the piston 11 also sucks some of the volume flow into the chamber 9b, from which it is then discharged in the range q>ι — φ₂. The operation is thus reverse to the compensation on the pressure side.

Figure 4 shows, in more detail, the pulsation S occurring in the volume flow Q of the fluid to be supplied as a function of the rotation angle φ, the pulsation being due to a delay caused by the compression of the unpressurized fluid. The fluid is first compressed, wherein the working piston 3 is not capable of producing a volume flow during this time, although it is in a stroke. The rotation angle φ_N> at which the pulsation occurs, depends, for example, on the total volume of the fluid to be compressed, its pressure, the fluid itself, and its compressibility, as well as the temperature. Figure 4 also shows the average volume flow Q , which substantially corresponds to the compensated volume flow Q of the outlet side P produced by the device shown in Fig. 1. Variation due to the compressibility is effective particularly in this device. At a rotational speed of 1500 rpm (r/min), there is a pulsation at the frequency of 50 Hz. In a periodically recurring volume flow, 360° (φ = 2π) corresponds to one revolution of the shaft and one reciprocating motion of the displacing means, i.e. the working piston. The figure also shows the stroke h of the auxiliary piston.

Figure 5 illustrates an advantageous embodiment of the invention applied in connection with the device of Fig. 1. The operation and the reference numbers in Figs. 1 and 5 correspond to each other. A pressure accumulator 22 is connected to the pressure side P via a choker 21. The function of the pressure accumulator 22 is, through the connection 23, to receive pressurized fluid under the working pressure from the pressure side P and to store it until it is needed again. The accumulator can be for example a weight-, spring- or gas-loaded accumulator, wherein it can be a piston accumulator, a bladder-type accumulator or a diaphragm accumulator. The accumulator is connected to the chamber 4 via a channel formed by the connections 23 and 24. The channel also comprises a valve 25, for example a check valve, to prevent a back flow of the fluid into the accumulator 22. Because of a small need of volume flow, the choker 21 can be very narrow, wherein the compressing volume flow is taken from the accumulator and not from the pressure side P. A particular advantage is that no pulsation is thus caused on the pressure side P and it is possible to use the volume flow and pressure of the system itself. Fluid is transferred from the accumulator 22, when the pressure of the chamber 4 is lower than the pressure of the accumulator 22 before the compression. Instead of the piston 3, the fluid under the working pressure and the pressure effect of the accumulator 22 compresses very fast the fluid under the suction pressure in the chamber 4, wherein the function of the piston 3 remains to supply fluid to the pressure side P via the pressure valve 7. Simultaneously with the compression and the increase in the pressure, the connection to the suction side T is closed. The compressing volume is thus compensated by a change in the volume of the accumulator 22 and by a supply of the fluid.

Pressure losses may occur in the channels 23 and 24, wherein to compensate for them, the piston 3 must compress the fluid to raise its pres- sure to the working pressure. However, when the fluid of the accumulator 22 participates in the work, the required rise in the pressure is considerably smaller, wherein the pulsation and the rotation angle φ_N are considerably smaller than in the prior art. The volume flow to be supplied from the accumulator 22 is also considerably small and has a short duration. During the suction phase, the channelling is closed so that the required fluid is not taken from the accumulator 22, disturbing the pressure side P, and so that a smaller pressure accumulator could be used. By means of a pressure loss caused by the choker 21 , it is possible to continuously charge the accumulator 22 from the pressure side P, but to simultaneously prevent the intake of fluid entering the working chamber 4 from the pressure side P instead of from the accumulator 22.

According to an advantageous embodiment of the invention, the channels 23 and 24 are closed and opened in a synchronized manner by the auxiliary piston 11. As can be seen from the Figs. 2a to 2d, the piston 11 can be used to control at least two working pistons 3 which are in reverse phases. In Fig. 5, the lower working piston 3 is in its lower dead point, wherein also temporary feeding takes place. Typically, the feeding period is substantially shorter than the time taken by the rotation angle Φ_N- The feeding is controlled by means of control edges formed in the auxiliary piston 11.

The control edges close or open the inlet channel 23 and the outlet channel 24 which are not connected to the auxiliary chamber 9. The piston 11 and the cylinder 10 are preferably designed in such a way that the pressure effect of the fluid under working pressure in the compressed volume 26 is compensated for and is not effective as a force resisting to the movement of the piston 11. At the same time, changes in the fluid pressure are prevented, particularly a decrease in the pres- sure. The volume 26 and the control edges can be formed, for example, by means of annular cuttings in the piston 11 and/or in the cylinder 10. In this context, it is possible to apply techniques known as such, used in connection with directional control valves equipped with movable gates. A more detailed application will be obvious as such for anyone skilled in the art, wherein a more detailed description will not be necessary.

In the present invention, the piston 11 forms a gate-like structure comprising control edges, to allow the movement and choking of the pressurized medium. In this way, a very simple, light-weight, integrated and compact structure is formed, combining several functions. The fluid supply into the chamber 4 is controlled by dimensioning and fitting the control edges in such a way that it takes place when the piston 11 is in a certain position/positions. Similarly, the closing is timed in a corresponding way. A change in the synchronization is made by changing the piston 11.

The supply takes place in both directions of motion of the piston 11 , wherein at least two working cylinders with different phases can be controlled. According to Fig. 5, the piston 11 controls two chambers 4, but the entry of the fluid into the chamber is controlled by means of the position of the opening 24c of the connection 24. Thus, the piston 3 opens and closes the opening 24c, which is opened at the lower dead point of the piston 3, wherein the opposite piston 3 is at the upper dead point, closing the opening 24c. The check valve can also be placed in the channel 24a and/or 24b. Each piston 11 can also control one or more chambers 4, or several control edges can be separately placed in one piston 11 for channels leading to different chambers 4. The second piston 11 can also be arranged in the way shown in Fig. 3, wherein only two auxiliary pistons are required for the damping and the compensa- tions. A very compact piston structure is obtained when the controlling piston 11 complies with that shown in Fig. 3, wherein the piston part is provided with corresponding control edges and the volume 26. In this case, the piston part must have a sufficient length to keep the channels 23 and 24 closed. It is obvious that control valves, such as cut-off valves, can be placed in the channels, to turn off the compensation.

With reference to Fig. 6, to control the timing in a more versatile and programmed way, if necessary, controlled valve means V are used instead of the piston 11 , such as an electrically controlled directional valve equipped with a gate, or cartridge valves. The control can thus be changed depending on the working pressure, rotational speed, the fluid, temperature, or another given parameter, wherein the operation of the engine can be considerably optimized. The required control system CTR comprises, for example, a computer or control logics which is equipped with memory means and a central processing unit and oper- ated under a control program, and which controls the valve means with electrical setting signals. The valves may also comprise a control card to which the control system is connected and which controls the operation of the valve more precisely. The control system CTR may be connected to sensor means, such as pressure, rpm, temperature and posi- tion sensors, to determine the state of the hydraulic engine. According to the stored setting parameters and the control algorithm contained in the computing program, the control system CTR determines the required control. The algorithm is used to implement the given optimization, timing, control or other adjustments. The invention is not limited solely to the above-presented embodiment, but it can be modified within the scope of the appended claims. For example, the pressurized medium can be water, aqueous emulsions and fluids, mineral oil and vegetable oil based hydraulic fluids and syn- thetic oils with varying compressibility. The type, position and number of check valves related to the channels may vary. The channels are preferably formed in the body of the engine. The invention can be applied in hydraulic engines which operate by different principles and are suitable for use as a pump and/or a motor. In an example, the body may rotate around a stationary shaft. As a mechanical solution, the invention is particularly suitable for radial engines, but the principle can also be applied in axial engines. The coupling can be implemented separately or preferably integrated in the hydraulic engine. The pressure accumulator can also be charged from a separate pressure source but, in the simplest way, from the pressure side.

Claims

Claims:

1. A coupling for equalizing variations in the volume flow of a hydraulic engine, the hydraulic engine (M) comprising at least one periodically operated displacement chamber (4) and being arranged for sucking pressurized medium into this displacement chamber (4) from the suction side (T) of the hydraulic engine, to compress the pressurized medium sucked into the displacement chamber (4) to a working pressure and to feed the compressed pressurized medium out of the displacement chamber (4) to the pressure side (P) of the hydraulic engine, characterized in that, to damp the variation caused by the compression, the coupling is arranged to supply pressurized medium into the chamber (4), to compress the pressurized medium sucked into this displacement chamber (4).

2. A coupling according to claim 1 , characterized in that the pressur- ized medium is arranged to be temporarily supplied to the displacement chamber (4), when the compression of the pressurized medium sucked into this displacement chamber (4) is being started in the hydraulic engine.

3. A coupling according to claim 1 or 2, characterized in that it com- prises a pressure accumulator (22) for storage and supply of pressurized medium, the pressure accumulator (22) being in contact with the displacement chamber (4) by means of controllable valve means (V).

4. A coupling according to claim 3, characterized in that the pressure accumulator (22) is connected to the pressure side (P) to maintain the pressure.

5. A coupling according to any of the claims 1 to 4, characterized in that the hydraulic engine (M) comprises a piston structure (11) synchronized with the operation of the displacement chamber (4) and arranged to control and time the supply of pressurized medium into the displacement chamber (4).

6. A coupling according to any of the claims 1 to 4, characterized in that the hydraulic engine (M) comprises at least one auxiliary displacement chamber (9, 9a) which is synchronized with the operation of the displacement chamber (4) and is connected to the pressure side (P), and which is, for equalizing the supply of the volume flow, arranged to receive some of the pressurized medium supplied to the pressure side (P) and to feed it back to the pressure side (P) by means of the piston structure (11), the piston structure (11) being arranged to control and time the supply of pressurized medium into the displacement chamber (4).

7. A coupling according to claim 5 or 6, characterized in that the piston structure (11) is arranged to close and open the channelling (23, 24) which leads to the displacement chamber (4) and which is arranged for the supply of pressurized medium.

8. A coupling according to any of the claims 1 to 7, characterized in that the hydraulic engine (M) comprises at least one auxiliary displacement chamber (9b) which is synchronized with the operation of the displacement chamber (4) and is connected to the suction side (T), and which is, for equalizing the supply of the volume flow, arranged, together with the displacement chamber (4), to suck pressurized medium from the suction side (T), when the sucked volume flow (Q) is smaller than a predetermined level, and to feed it back to the suction side (T) by means of the piston structure (11), when the sucked volume flow (Q) is greater than said level.

9. A method for equalizing variations in the volume flow of a hydraulic engine, the hydraulic engine (M) comprising at least one periodically operated displacement chamber (4), in which method pressurized medium is sucked into this displacement chamber (4) from the suction side (T) of the hydraulic engine, the pressurized medium sucked into the displacement chamber (4) is compressed to a working pressure, and the compressed pressurized medium is fed out of the displacement chamber (4) to the pressure side (P) of the hydraulic engine, characterized in that, to damp the variation caused by the compression, pressurized medium is supplied into the displacement chamber (4) to compress the pressurized medium sucked into this displacement chamber (4).

10. A method according to claim 9, characterized in that the pressur- ized medium is temporarily supplied to the displacement chamber (4), when the compression of the pressurized medium sucked into this displacement chamber (4) is being started in the hydraulic engine.

11. A method according to claim 9 or 10, characterized in that the hydraulic engine (M) comprises a piston structure (11) synchronized with the operation of the displacement chamber (4) to control and time the supply of pressurized medium into the displacement chamber (4).

12. A method according to claim 9 or 10, characterized in that the hydraulic engine (M) comprises at least one auxiliary displacement chamber (9) which is synchronized with the operation of the displace- ment chamber (4) and is connected to the pressure side (P), and which is, for equalizing the supply of the volume flow, arranged to receive some of the pressurized medium suppied to the pressure side (P) and to feed it back to the pressure side (P) by means of the piston structure (11), the piston structure (11) being arranged to control and time the supply of pressurized medium into the displacement chamber (4).