FI110960B - Connection and method for smoothing volumetric flow variations in a hydraulic machine - Google Patents

Description

1,110,960

Connection and method for smoothing volumetric flow variations in a hydraulic machine

The invention relates to a coupling according to the preamble of claim 1 for equalizing the volumetric flow variation of a hydraulic machine. The invention also relates to a method according to the preamble of claim 9 for smoothing the volumetric flow variation of a hydraulic machine.

Hydraulic machines operating on the displacement principle are high in power density because they are suitable for very high pressures, typically up to 300-400 bar and above. In a hydraulic pump, mechanical energy, such as torque and rotation speed, is converted into hydraulic energy, i.e. pressure of the pressure medium and its volume flow. In the hydraulic raul motor, the hydraulic energy is again converted into mechanical energy. Rotational speeds are typically 500-10000 rpm. The torque given by the hydraulic motor depends on the engine speed and the pressure difference across it. Power still depends on the volume flow and the pressure difference across the engine. Slow-rotating machines, typically at speeds between 0.5 and 1000 rpm, are designed to deliver high torque at low speeds. These machines are called LSHT (Low Speed High Torque) machines. Many hydraulic machines can act as both pumps and motors due to their principle of operation. Pumps operating on the displacement principle have mechanically sealed displacement chambers in which the fluid is sucked, for example, by the movement of the piston on the suction side or the inlet side, and transported to the pressure side 25 or to the outlet by the movement of the piston. In the case of motors, the opposite is true.

A large drop in pressure in the chambers or an increase in pressure will cause shocks to the machine frame and hence frame sounds which will be audible as ambient noise emitting into the environment. In addition, the noise 30 is caused by the volumetric flow variation due to the uneven distribution of the above volumes relative to the rotation angle of the machine. These volumes are formed, for example, in an axial or radial piston pump wood or the like. Pressure variations also lead to resonances in other system-related structures, which further add. 35 are a major source of noise.

4,210,960

German Application Publication No. 1703210 discloses one solution for reducing variations in the volume flow produced by a hydraulic machine, whereby a very small or even number of pistons can be used. An example is a two-piston axial piston pump. Volume-5 current fluctuations, on the other hand, can be reduced by adding pistons, typically 7 to 11, which, however, results in a larger, more complex, and more expensive structure. Typically, an odd number of pistons is also used, which achieves an equal volume flow rate as doubled but an even number of pistons.

The solution disclosed in application publication 1703210 comprises a pair of synchronously operating auxiliary pistons which supply a volumetric flow to the output side when the volumetric flow supplied by the working cylinders is below a specified average. The radial auxiliary pistons receive a volume flow from the outlet side when this volume flow is fed to the working cylinders in excess of a specified average. Fluctuations in volumetric flow depend on the rotation angle of the drive shaft, whereby the auxiliary pistons are coupled to the drive shaft for synchronization. The displacement volume required by the auxiliary pistons is considerably smaller compared to the working cylinders. This makes it easy to place them in the machine, for example, the four-piston machine 20 needs only one auxiliary piston. However, due to the balance of power, two opposite auxiliary pistons are preferable.

Alternatively, when the hydraulic machine acts as a motor, the auxiliary piston volume is connected to the pressurized inlet by means of controllable valve means, for example a 3/2-way valve. In the motor drive 25, the auxiliary pistons can compensate for variations in output torque. The auxiliary piston driven by the pressurized pressure medium provides the required additional torque at the desired moment of movement of the slave cylinder. To remove the pressure medium, the auxiliary piston instead uses a torque force, so the auxiliary piston smoothes out the torque fluctuations. The friction varies in different directions of motion of the auxiliary piston, but causes a relatively small variation in torque compared to the case where the auxiliary piston is not used for the compensation described above.

The problem is that the functioning of the hydraulic pump suction-side volume flow also takes place in the pressure and volume flow fluctuations, 35 which depend on the rotation angle. In addition, the pressurized volume of the outlet side 3 110960 flow will pulsation caused by the compression of the pressure medium, in particular at the time when pressurized, liquid is sucked into the displacement chamber will compress the load, for example, against a pressure valve, wherein the pressure increases at the same time. The pressure is increased to a working operating pressure that can be controlled by a pressure regulator connected to the hydraulic machine or by a pressure relief valve connected to the system. The pressure in the valve is typically exerted on a movable, flexible socket or slide, whereby adjusting the spring tension can change the maximum pressure. In hydrostatics, external force is known to exert a pressure on the medium, which is propagated in all directions in the liquid. The magnitude of the pressure is obtained by dividing the force by the effective surface area.

However, insoluble air entrained in the fluid causes the fluid to compress as the pressure increases, thereby reducing the volume of the fluid. Rapid pressure fluctuations thus cause pressure shocks to the structures due to volume changes. Compression is described as a pressure medium dependent constant, for example about 0.7-0.8% for mineral oil and about 0.45% for water for 100 bar. In the hydraulic pump, the fluid must be compressed before the pressure advance is reached and before sufficient working pressure is reached and the volume flow is transferred to the pressure side. The larger the compressible volume, the longer it takes to reach working pressure and the greater the disruption.

The object of the present invention is to eliminate the above-mentioned problems. The purpose is to introduce a circuit that can compensate for the delay due to compression. The circuit according to the invention is characterized by what is disclosed in the characterizing part of claim 1. The method according to the invention is characterized by what is set forth in the characterizing part of claim 9.

The invention is based on the idea that a displacement chamber is supplied from a pressure source, such as a pressure accumulator, by a substantially pressurized pressure medium. Management occurs just before the chamber, the suction-side pressure of the pressure medium begins to compress, for example, by means of the working piston. By supplying a volume accumulator of pressure accumulator, the medium-35 is rapidly brought to the working pressure, so that the compression 110960 is not used solely by the piston, and the medium can be previously supplied at working pressure out of the displacement volume. In this case, the variations are significantly reduced compared to the prior art.

According to a preferred embodiment of the invention, the pressure accumulator 5 is charged through the throttle from the pressure side of the hydraulic machine at working pressure and the supply of medium to the chamber is controlled mechanically or electrically by means of valves. According to a preferred embodiment of the invention, the hydraulic machine comprises auxiliary pistons for the compensation of the volume flow whose movement is utilized to close and open the connection between the pressure accumulator and the volume. The auxiliary pistons then act as said valve means. The operation is based on the fact that the fluid supply from the pressure accumulator can be synchronized with the movement of the auxiliary piston. The auxiliary piston can also act as a mere moving valve slide.

According to a preferred embodiment of the invention, the auxiliary pistons 15 on the suction side is used to reduce the volume flow variations.

These variations are similarly dependent on the angle of rotation, since the same working pistons are used to absorb the compressible pressure medium. At the same time, pressure variations that cause cavitation and machine wear, as well as structural resonances and noise due to variations are reduced.

Up to twice the number of pistons is required to provide compensation, i.e. auxiliary pistons, for both the suction side and the pressure side. The second pair of cross-reciprocal auxiliary piston functions with the auxiliary piston pair. This makes the placement and preparation of the machine much more difficult, leading to larger and heavier equipment. To eliminate the problem presents a new piston design of suction pressure side of the volume flow to compensate for fluctuations. A major advantage of the compensation piston is space saving, weight control and integrated design. The principle is to form two displacement chambers in the same auxiliary piston, whereby the piston moves within the cylinder volume while forming two separate chambers which operate in opposite steps.

The medium supply and the output-side compensation is achieved by using a single substantially steady volume flow. Compression 5 110960 removes residual pulsation from the compensated volume flow. The input and output side of the compensation is achieved by a low-noise hydraulic machinery.

The invention will now be explained in more detail by way of a preferred embodiment, with reference to the accompanying drawings, in which Figure 1 shows a prior art 2-piston pump comprising two compensation pistons, Figures 2a to 2d illustrate the operation of the compensation pistons Fig. 4 is a schematic diagram of the flow rate variation of the 2-piston pump and the pulsation caused by the compression, Fig. 5 shows a 2-piston pump of the preferred embodiment, comprising a connection for venting the pulse, and illustrates an alternative preferred embodiment 20 for the coupling of FIG.

Fig. 1 is a simplified representation of a hydraulic machine M known per se, which operates as a pump on the displacement principle and comprises two additional pistons for volume flow compensation. The pump is a 2-piston radial piston pump driven by shaft 1.

A shaft 1 is provided with an eccentric 2 along which the piston 3 slides. Two displacement chambers 4 are fitted to the pump body B by means of cylinders 5, in which cylinder 5 the piston 3 moves reciprocally sealed. The chamber 4 communicates with the suction side T for suction of the liquid pressure medium by the vacuum to the chamber 4 which expands the movement of the piston 3; 30 thanks. During suction, suction valve 6 is open and pressure valve 7 6 110960 is closed. At the lower dead center of the piston 3, the suction changes to compress, whereby the fluid pressure rises substantially to the working pressure and is supplied to the pressure side P. During the supply, the suction valve 6 is closed and the pressure valve 7 is open. The working pressure and the pressure side P-5 to the maximum pressure may be adjusted, for example, the pressure control valve 8 as a pressure valve. Other loads connected to the pressure side P may also determine the working pressure.

At the upper dead center of the piston 3, the feed changes again for suction. The shaft I is rotated, for example, by an electric motor, whereby the pump operates periodically, sinusoidally. The opposite pistons 3 are arranged to operate in the opposite phase, whereby one of the suctiones the other to supply liquid. The eccentric 2 of Figure 1 is arranged such that each piston 3 performs one suction and supply during a rotation of the shaft 1.

Figure 1 also shows two prior art displacement chambers 9 formed by a cylinder 10 in which the piston II is reciprocally sealed. The piston 11 is guided by a cam 12 mounted on the shaft 1, along which the bearing 13 of the piston 11 slides. The cam 12 is arranged to move the piston 11 back and forth twice within one revolution of the shaft. The displacement volume of the piston 11 is about 21% of the displacement volume of the piston 3 to equalize the volume flow Q. Correspondingly, with a 4-piston pump, the volume is only about 4%, whereby the pistons 11 perform four strokes per revolution. The auxiliary chamber 9 is synchronized to receive a volume flow from the working chamber 4 to the pressure side P via ducting 28 when the common supply of the working chambers 4 is higher than the average or selected level. The auxiliary chamber 9 is synchronized to deliver a volume flow from the auxiliary chamber 9 back to the pressure side P when the common supply of the working chambers 4 is below the average or selected level. The profile of the cam 12 can mechanically schedule operation and control the movement of the piston 30 to control the volume flow. Two opposing auxiliary pistons * 11 move at the same rate.

Fig. 2a is a more detailed diagram of the stroke h of the two working pistons 3 as a function of the rotation angle φ of the axis 1, where φ = 2π corresponds to one rotation. Fig. 2b shows one pressure side of the working pistons 3; The variation of the volume flow Q produced by 35 P as a function of the rotation angle josta, of which 7 110960 is seen to be greater than the mean QM between <Pi-cp2 and lower between φ2-φ3 and φ0-cpi. The same is repeated every half cycle. The auxiliary pistons 11 receive from the pressure side P between tilι and Φ2 an additional volumetric flow marked by oblique lines and deliver it to the pressure side P at subsequent intervals <p2-cp3 and φ0-Φτ. Fig. 2c shows the position of the auxiliary pistons 11 as a function of the rotation angle cp. Fig. 2d shows the stroke h of the auxiliary pistons 11 as a function of the rotation angle φ of the axis 1, which is struck and synchronized by the shaped cam 12.

10 When the hydraulic engine operates as a motor, the input side of the T compensation circuit is implemented as a multiplexer in Figure 3 27a and 27b, but the multiplexer 27 is connected to the chamber of Figure 3 9a or the chamber of Figure 1 9. It is understood that the channels may be positioned control valves, such as gate valves and non-return valve, whereby the operation does not depend on the suction valve 6 or the positioning of the ducts 15 and are electrically controlled. It must be possible to switch off the compensation, especially when the inlet side T acts as the suction side of the pump.

The pump inlet side of the T coupling to compensate for the volume flow is shown in simplified form in Figure 3, wherein a dual-chamber 20 is applied to the auxiliary piston 11. The auxiliary piston 11 is a function of Figures 2a to 2d. Alternatively, each auxiliary piston 11 takes care of separately either the output side or the input side of the compensation to the following description, only one of the chambers 9a, 9b is in use. At least two separate displacement chambers 9a and 9b are formed on the hydraulic machine for the same stroke.

* 4. ·.

In the example, the piston 11 provided with the shaft portion and the piston portion moves in a cylinder 10 formed with a chamber in which the shaft portion moves in a sealed manner. The piston portion attached to the shaft portion divides the chamber into two displacement chambers, so that as the chamber 9a expands during suction, the chamber 9b decreases during feeding and vice versa. It may also be driven that the plunger 11 comprises a plurality of plunger sections adapted to move simultaneously in two or more separated chambers. The suction side T can be offset by a pair of auxiliary pistons which, with reference to Figure 1, are guided by the same shaft 1 and the same cam 12 having a substantially oval cross-section. The other auxiliary pistons 35 are then at an angle of rotation of 90 ° to the auxiliary pistons 11 8 110960, the phase being opposite. As the cam 12 guides the auxiliary pistons 11 to four strokes during a turn, the second auxiliary pistons must be positioned at a 45 ° angle so that the cam 12 is substantially a square whose corners are strongly rounded and the sides are concave. The chamber 9b can also be provided with a crankcase of the hydraulic machine M. This gives the advantage that the volume change of the chamber 9a and the crankcase are equal, so that the volume flows controlled by the chambers 9a, 9b are also equal.

Referring to Figure 3, chamber 9b communicates with suction side 10 via ducting 27 with one piston 11 and chamber 9b controlling at least two piston 3s of different stages, whereby ducts 27a and 27b are led to inlet side before suction valves 6. In this case, suction valve 6 and the other is opened as a result of the suction step, wherein the volume flow delivered by the chamber 9b 15 is directed to the suction piston 3 as shown in Fig. 2b between <p1 - cp2. The suction phase is depicted by a curve 3 in Fig. 2a and the change in volume flow Q at suction side T is shown in Fig. 2b. The movement of the piston 11 is shown in Fig. 2c, whereby the piston 3 receives a more volumetric flow according to the curve QM. The pressure changes are reduced and, in particular, the occurrence of cavitation is avoided so that the pressure cannot fall to the vaporization pressure. When the suction volume flow is smaller than the mean Qm between cpo - q> i, φ2 - q> 3, the piston 11 also sucks a volume flow into the chamber 9b, from which it is then discharged between välillä - φ2. The operation is thus the opposite side compared to the pressure compensation.

Figure 4 illustrates in more detail the pulsation S in the volume flow Q of the fluid introduced as a function of the rotation angle φ due to the delay due to the compression of the unpressurized fluid. The liquid is first compressed, so that during this time the piston 3 cannot produce a volume flow even though it is in stroke motion. The angle of rotation φΝ, during which the pulsation occurs, depends, e.g. the total volume of the fluid to be compressed, the pressure, the fluid itself and its compressibility and temperature. Figure 4 also shows the average volume flow Q M which substantially corresponds to the apparatus shown in Figure 1 to provide the compensated output from the P side of the volume flow Q

35 Variability due to compressibility is particularly affecting this unit. At a speed of 1500 rpm (rpm), the pulse rate emerges at 9 110960 at 50 Hz. In a periodic flow of volume, 360Q (φ = 2π) corresponds to one rotation of the shaft and one reciprocating motion of the displacement member, the piston. Also shown is the movement h of the auxiliary piston.

Figure 5 illustrates a preferred embodiment of the invention when applied to the apparatus of Figure 1. The operation and reference numerals in Figures 1 and 5 correspond to each other. The pressure accumulator 22 is connected to the pressure side P via the choke 21 to receive and store (until needed again) the pressurized, pressurized fluid from the pressure side P. The accumulator 10 may be, for example, a weight, spring or gas it may be a piston, bladder or diaphragm battery The battery is connected to the chamber 4 via the channels 23 and 24.

The channel also has a valve 25, for example a non-return valve, to prevent the liquid from returning to the accumulator 22. Due to the need for a small volume flow 15, the throttle 21 can be very narrow, whereby the compressible volume flow is taken from the battery rather than from the pressure side P. This has the particular advantage that it does not cause pulsation to pressure side P and system volume flow and pressure. Liquid is transferred from the battery 22 when the pressure in the chamber 4 is lower than the pressure of the battery 22 from the compressor-20 before. Instead of the piston 3, the pressurized fluid and the pressure action of the accumulator 22 compresses the suction fluid in the chamber 4 very rapidly, leaving the piston 3 to supply the liquid to the pressure side P via the pressure valve 7. Simultaneously with the compression and the increase in pressure, the connection to the suction side T is closed. The compressible volume of 25 is thus compensated by a change in the volume of the battery 22 and by the supply of fluid.

In the ducts 23 and 24, pressure drops can occur, so that the piston 3 must be pressurized with liquid to compensate for them to bring it to working pressure. However, when the fluid 22 of the battery 22 is involved in the work, the increase in pressure 30 required is much smaller, whereby the pulsation and rotation angle cpN are significantly lower than the prior art. The volume flow supplied from the battery 22 is also remarkably small and short-lived. During the suction phase, the ducting is closed so that the required fluid is not withdrawn from the battery 22, interfering with the pressure side P, and a miniature-35-size pressure battery can be used. By means of the pressure drop caused by the choke 21, the accumulator 22 can be continuously charged from the pressure side P, but at the same time the fluid entering the work chamber 4 is prevented from being drawn from the pressure side P instead of the accumulator 22.

According to a preferred embodiment of the invention, the channeling 23 5 and 24 is synchronously closed and opened by the auxiliary piston 11. As shown in Figures 2a to 2d, the piston 11 can be controlled by at least two working pistons 3 which are in the opposite phase. In Fig. 5, the lower working piston 3 is at its lower dead center, whereby also a short-term supply occurs. Typically, the feed time is substantially shorter than the rotation time φΝ. The feed is guided by guide edges formed in the auxiliary piston 11.

The guide edges close or open the inlet 23 and the outlet 24 not communicating with the auxiliary chamber 9. The piston 11 and cylinder 10 are preferably configured to compensate for the pressure effect of the pressurized fluid 15 in the sealed volume 26 and not act as At the same time, changes in the fluid pressure, in particular the pressure drop, are prevented. The volume 26 and the guide edges can be formed, for example, by annular turning in the piston 11 and / or in the cylinder 10. In this connection, 20 techniques known per se and utilized in connection with directional valves with movable slides can be applied. As will be apparent from the foregoing, more detailed application will be apparent to those skilled in the art, so no further explanation is required.

In the present invention, the piston 11 forms a skid-like structure and includes guide edges 25 to allow and throttle the passage of pressure medium. This creates a very simple, lightweight, integrated and compact structure that combines multiple functions. The fluid supply to the chamber 4 is controlled by dimensioning and adjusting the guide edges such that the piston 11 is in a certain position or position 30. Similarly, the closure is timed accordingly. The timing change is done by changing the piston 11.

The feed takes place in both directions of movement of the piston 11, so that at least two different working cylinders can be controlled. As shown in Figure 5, the piston 11 controls the two chambers 4, but the fluid entry into the chamber m 11 110960 is controlled by the location of the orifice 24c of the connection 24. The piston 3 then opens and closes the orifice 24c, which is open at the lower dead center of the piston 3, whereby the opposite piston 3 is at the upper dead center, closing the orifice 24c. Counter-valves may also be provided in the passageway 5 24a and / or 24b. Each plunger 11 may also guide one or more chambers 4, or a plurality of guide edges may be disposed separately in one plunger 11 for passageways to the various chambers 4. The second piston 11 may also be arranged as shown in Figure 3, whereby only two auxiliary pistons are required for damping and compensation. A very compact piston structure 10 is obtained when the guiding piston 11 is in accordance with Figure 3, whereby the piston member is provided with corresponding guide edges and a volume 26. Thus, the piston member is of sufficient length to keep the channels 23 and 24 closed. It will be appreciated that control valves, such as shut-off valves, may be provided in the ducts, whereby the compensation may also be deactivated.

Referring to Fig. 6, controlled valve means V, for example an electrically controlled slide valve or cartridge valves, are used instead of the piston 11 to control the timing in a more versatile and programmed manner. The control can thus be varied 20 depending on operating pressure, rotational speed, fluid, temperature or other desired parameter, whereby the operation of the machine can be optimized considerably. The required control system CTR comprises, for example, a control program computer or control logic 25 having memory means and a central processing unit which controls the valve means by means of electrical setting signals. Valves: may also comprise a control card to which the control system is connected and which more precisely controls the operation of the valve. The control system CTR may communicate with sensor means such as pressure, speed, temperature and position sensors to determine the status of the hydraulic machine. In accordance with the stored control parameters and the control algorithm contained in the calculation program, the control system CTR determines the control required. The algorithm performs the desired optimization, timing, control or other adjustment.

The invention is not limited to the above embodiment only, but can be modified within the scope of the appended claims. For example, the pressurized medium may be water, water-based emulsions and fluids, mineral oil and vegetable oil-based hydraulic fluids, and synthetic oils of varying compressibility. The type, location and number of check valves associated with ductwork may vary. The ducts are preferably formed on the machine frame. The invention can be applied to hydraulic machines operating on different principles, which are suitable for operation as a pump and / or motor. In the example, the frame may rotate about a fixed axis. As a mechanical solution, the invention is particularly applicable to radial machines, but the principle can also be applied to axial machines. The coupling may be performed separately or preferably 10 integrated with a hydraulic machine. A pressure accumulator can also be charged from a separate pressure source, but in the simplest way from the pressure side.

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Claims (12)

A coupling for equalizing variation in the volume flow of a hydraulic machine, said hydraulic machine (M) comprising at least one cyclically functioning thrust chamber (4), and which is arranged to suck pressure medium into said thrust chamber (4) from the suction side (T). ) of the hydraulic machine, to compress the pressurized chamber (4) to suck the pressure medium into a working pressure, and to feed the compressed pressure medium out of the pressure chamber (4) to the pressure side (P) of the hydraulic machine, characterized in that The coupling is adapted to supply the pressure medium to the chamber (4) in order to compress the pressure medium to this pressure chamber to absorb the variation caused by the compression (4). Coupling according to claim 1, characterized in that the pressurized medium 15 is arranged to be measured momentarily in this pressure chamber (4) at the beginning of compression of the pressure medium sucked into this pressure chamber (4) with the hydraulic machine. Coupling according to claim 1 or 2, characterized in that it comprises a pressure accumulator (22) for storing and feeding pressure medium under pressure, which pressure accumulator (22) starts in communication with the pressure chamber (4) by 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. Coupling according to any of claims 1-4, characterized in that the hydraulic machine (M) comprises a piston structure (11) synchronized with the function of the thrust chamber (4), which is arranged to control and adjust the input of the pressure medium under pressure to the thrust chamber. (4). Coupling according to any of claims 1-4, characterized in that the hydraulic machine (M) comprises at least one synchronized with the function of the thrust chamber (4), in connection with the auxiliary thrust chamber (9, 9a) existing in the pressure side (P). , which is, for smoothing the input of the volume flow, arranged to receive a portion of the pressure medium fed to the pressure side (P) and to feed it back to the pressure side (P) by means of a piston structure (11), which piston structure 5 (11) is arranged to control and adjust the input of pressure medium. under pressure to the squeeze chamber (4). Coupling according to claim 5 or 6, characterized in that the piston structure (11) is arranged to close and open a ducting (23, 24) leading to the thrust chamber (4), which is arranged for supplying pressure medium under pressure. Coupling according to any of claims 1-7, characterized in that the hydraulic machine (M) comprises at least one auxiliary penetration chamber (9b) synchronized with the function of the thrust chamber (4), which is in connection with the suction side (T) , in order to equalize the suction volume flow, arranged to suction pressure medium from the suction side (T) together with the extraction chamber (4), when the suction volume flow (Q) is less than a certain level, and to feed it back to the suction side (T) with by means of a piston structure (11), when the suctioned volume flow (Q) is higher than said level. 20 9. A method for equalizing variation in the volume flow of a hydraulic machine, said hydraulic machine (M) comprising at least one cyclically functioning thrust chamber (4), and in which method pressure medium is sucked into said thrust chamber (4) from the suction side (T). of the hydraulic machine, the pressure medium 25 sucked into the pressure chamber (4) is compressed into a working pressure, and the compressed pressure medium is measured out of the pressure chamber (4) at the pressure side (P) of the hydraulic machine, characterized in that in order to dampen the pressure caused by The pressure medium variation under pressure is measured in the pressure chamber (4) to compress the pressure medium sucked into this pressure chamber (4). Method according to claim 9, characterized in that pressure medium under pressure is measured momentarily in the pressure chamber (4) at the beginning of compression of the pressure medium sucked into the pressure chamber (4) with the hydraulic machine. 18 110960 Method according to claim 9 or 10, characterized in that the hydraulic machine (M) comprises a piston structure (11) synchronized with the function of the thrust chamber (4), which is arranged to control and adjust the input of the pressure medium under pressure to the thrust chamber ( 4). Method according to claim 9 or 10, characterized in that the hydraulic machine (M) comprises at least one auxiliary penetration chamber (9) synchronized with the function of the thrust chamber (4), which is in order to equalize the input. of the volume flow, arranged to receive a portion of the pressure medium fed to the pressure side (P) and to feed it back to the pressure side (P) by means of a piston structure (11), which piston structure (11) is also used to control and adjust the input of the pressurized medium under pressure into the thrust chamber (4).