The invention relates to a hydraulic machine having a gearwheel, which is arranged on a shaft so that they rotate together and has a first external toothing, a gear ring, which has a second external toothing and a first internal toothing that engages with the first external toothing, a toothed rim, which has a second internal toothing that engages with the second external toothing, and having a commutation arrangement for supplying and discharging in the correct order hydraulic fluid in pressure pockets which are formed in the first and/or second toothed coupling in each case between external and internal toothing.

Machines of that kind have the advantage that they run relatively slowly whilst providing high torque. They are described, for example, in US 2 989 951, EP 0 367 046 Bl or EP 0 038 482 A2.

In the known cases, the commutation may be difficult, because the pressure pockets move with respect to the connections, but supplying of the pressure pockets in the correct order is ensured in different ways. All constructions share the disadvantage, however, that the gear assembly comprising gearwheel, gear ring and/or toothed rim is not uniformly loaded. The pressure of the hydraulic fluid on the gear assembly is not uniform. In simplified terms, the high pressure spreads out in a half-moon shape which extends for instance over half of the gear assembly or a control plate. The pressure on the other half of the gear assembly (both halves viewed circumferentially) is low pressure (tank pressure) . These differences in

pressure lead to a tilting moment on the control plate. The gear ring in particular is affected by this tilting moment, and can easily cant between the gear wheel and the toothed rim from an inherently desirable even alignment. This results in uneven abrasion of the toothed couplings involved, that is, to increased wear and tear. Moreover, internal leakage can increase, which has an adverse effect on efficiency.

The invention is based on the problem of improving operational behaviour in a machine of that kind.

In a hydraulic machine of the kind mentioned in the introduction, that problem is solved in that on both axial sides of gearwheel and gear ring there is a lateral plate, which plates are joined axially to one another and move jointly with the gearwheel or the gear ring.

In this manner a closed unit is obtained, in which the gear assembly is housed. The two lateral plates, which are joined axially to one another, accommodate gearwheel and gear ring between them and support both parts similarly. Because the lateral plates move with the gearwheel or gear ring, the relative movement between the other of the two parts and the two lateral plates is not eliminated, but reduced to a relatively small value. At least between one of the two parts and the lateral plates there is no relative movement at all. The lateral plates and the gear assembly are together, as a unit, more stable than the individual parts of the gear assembly by themselves. Non-uniform force conditions can also be better compensated in this manner. The operational behaviour of the machine is improved, because leakage can be kept lower and wear and tear, at least in the region of the toothings, also

remains low. A hydraulic equilibrium occurs, because the same pressure is generated and maintained on both sides of the gear assembly. A tilting moment is therefore avoided.

In a preferred embodiment, provision is made for each lateral plate to have a through-opening for each pressure pocket, one of the lateral plates being covered by a rotary slide valve having inlet and outlet openings, and the other being covered by an end plate. The lateral plate adjacent to the rotary slide valve is therefore used to provide a duct arrangement for the necessary commutation of the hydraulic fluid to the pressure pockets. The ducts are in that case formed by the through-opening for each pressure pocket. Here, a pressure now builds up between the rotary slide valve and the lateral plate. On the opposite side of the gear assembly, however, there is arranged a similar lateral plate with corresponding through-openings, so that the corresponding pressure also builds up on the opposite side of the unit. The same pressure conditions are therefore present on both sides of this unit comprising lateral plates and gear assembly. It is therefore virtually impossible for a tilting moment to occur.

For each pressure pocket a pressure compensation duct is preferably provided between the two lateral plates. The desired pressure compensation can therefore be ensured also between the two lateral plates. Provided that between the teeth of the external toothing on the gearwheel and the teeth of the internal toothing on the gear ring there are still gaps, which form the pressure pockets, it is clear that the hydraulic fluid can convey the pressure through this opening so that the hydraulic pressure is the same on both sides of the

gear assembly. There is always a position, however, in which a tooth of the external toothing of the gearwheel virtually completely fills a tooth space in the internal toothing of the gear ring. In that case, pressure compensation could not take place without the pressure compensation duct, and there would again be a pressure imbalance between the two axial sides of the gear assembly. This imbalance is prevented, however, by the pressure compensation duct. No appreciable leakage is associated therewith. The duct merely ensures that the pressure from one axial side of the gear assembly is able to reach to the other side.

In an especially preferred construction, provision is made for the pressure compensation duct to be in the form of a groove in the base of an inter-tooth space. No further measures for control of the fluid then have to be taken. Pressure compensation is effected automatically. The paths are kept short.

On its side facing the lateral plate the rotary slide valve preferably has a series of control openings, which are arranged in a circle and are alternately connected to an inlet and an outlet connection. In this case the number of control openings differs from the number of openings in the lateral plate, to be precise, such that on a relative movement of lateral plate with respect to the rotary slide valve or vice versa the pressure pockets are supplied with hydraulic fluid always in the correct order. For example, with eleven pressure pockets there are twelve control openings in the rotary slide valve that are connected to the inlet connection and twelve control openings that are connected to the outlet connection. In this manner, supply of fluid to the pressure pockets in the

correct order can be ensured even when both the gear ring and the gearwheel rotate.

In that case it is especially preferred for the control openings to be connected alternately to one of two annular grooves on the axially opposing side of the rotary slide valve. The annular grooves ensure supply to the rotary slide valve in a simple manner. The rotary slide valve can then be permanently connected to the inlet and to the outlet.

In a very much preferred construction, provision is made for auxiliary openings to be arranged between each of the openings in the lateral plateε. Since the control openings in the rotary slide valve are always connected alternately to the inlet and the outlet, a high pressure always alternates with a low pressure at the rotary slide valve in the control openings. This alternately high and low pressure is now also transmitted to the lateral plate through the auxiliary openings, so that a certain pressure compensation over the surface is effected thereby.

The construction becomes quite especially advantageous when corresponding auxiliary openings in the two lateral plates are connected to one another. The pressure compensation which takes place on one surface can then also extend to the other lateral plate, so that both an axial equilibrium of forces over the unit comprising the lateral plates and the gear assembly and an equilibrium based extensively on superficial area is achieved.

The auxiliary openings are advantageously connected to holes through which axially running retaining bolts pass; the boltε connect the two lateral plates axially

to one another. Not even additional ducts are therefore needed to provide the pressure compensation between the auxiliary openings. On the contrary, the pressure compensation can be effected along the retaining bolts, which can be in the form, for example, of screw bolts. The construction thus becomes structurally relatively simple.

The retaining bolts preferably pass through the gearwheel. On the one hand, this means that the lateral plates rotate synchronously with the gearwheel. On the other hand, the axial retaining force is produced over a relatively small radius so that no great lever arms can form. The retaining bolts surround, of course, the shaft with which the gearwheel rotates. The retaining bolts in that case pass through the gearwheel as far towards the outside as possible, so that the axial connection between the lateral plates is effected relatively close to the pressure pockets, that is, where pressures can develop which, without the retaining bolts, would lead to forcing apart of the two lateral plates.

In a preferred construction, the lateral plates rotate synchronously with the gearwheel and the rotary slide valve rotates synchronously with the gear ring, a geared arrangement being provided which, of the rotating and the orbiting movements of the gear ring, transmits only the rotating movement to the rotary slide valve. In this manner, the relative movements that are obtained between the rotary slide valve and the lateral plate are exactly the same as those between the gear ring and the gearwheel. This facilitates supply in the correct order to the individual pressure pockets, that is the commutation. The rotary slide valve should perform only a rotating movement, whereas

the gear ring performs simultaneously an orbiting and a rotating movement. Both movements need not, however, be performed in the same direction. The geared arrangement acts therefore as a kind of filter which filters out a certain movement.

In that case it is especially preferred for the geared arrangement to be in the form of a sleeve which is connected to the gear ring and which has several recesses distributed circumferentially into which a corresponding number of projections of the rotary slide valve project. The sleeve here has a larger diameter than the rotary slide valve. It is not necessary for all projections of the rotary slide valve to be in operative engagement with all recesses at the same time. Basically, this operative engagement is necessary only between one projection and one recess. Generally, however, several projections will co-operate simultaneously with their corresponding recesses, so that the rotary movement is transmitted to the rotary slide valve whilst the orbiting movement of the gear ring leads to a relative displacement between the projections and the recesses.

The recesses are here preferably arranged in a projection running round the internal circumference of the sleeve at the end thereof. Even with a comparatively thin wall thickness of the sleeve, the necessary depths of the recesses can be obtained in this manner so that a relatively large eccentricity in the movement of the gear ring compared with the gearwheel is allowed. This eccentricity iε the reason for the displacements between the rotary slide valve and the sleeve in a radial direction.

At its external toothing the gear ring preferably has at least two sections of teeth having a greater thickness than the remaining teeth, the sleeve being pushed with corresponding cut-out areas onto the teeth of enlarged thickness. Because the sleeve is arranged on the teeth of the external toothing, there is on the one hand a large lever arm available for transmission of torque from the gear ring to the sleeve. Since, on the other hand, the teeth are used, manufacture is relatively simple. The gear ring can be manufactured according to known manufacturing techniques, for example, sintering. To make the projections with which the sleeve engages, it is merely necessary to grind the gear ring down, at least in the region of its teeth on the parts of the circumference not needed, that is to say, to reduce the thickness of the teeth.

On its side remote from the associated lateral plate, the end plate is advantageously of spherical form and bears against a correspondingly shaped housing wall. In this manner it is possible to compensate for defects that could occur on deformation of the shaft, for example, when the shaft is heavily loaded on one side. The rotating parts as a whole are then able to tilt a little. So that this does not result in canting forces, the end plate is of spherical construction and is able to find its own position in the housing. Even if there is lack of planarity during start-up, the end plate can bring itself into the correct position and then close a leak resulting from the lack of planarity.

The shaft is preferably supported on both sides of the gearwheel. This also allows relatively large lop-sided forces to act on the shaft. In principle it would then be sufficient to use a relatively weak bearing on one

side if the support is effected at a relatively large distance from the gearwheel.

The invention is described hereinafter with reference to a preferred embodiment in conjunction with the drawings, in which:

Fig . 1 is a cross-section through a hydraulic motor,

Fig . 2 is a diagrammatic perspective view of a gear assembly,

Fig . 3 showε a lateral plate,

Fig . 4 shows a rotary slide valve,

Fig . 5 shows a piston,

Fig . 6 shows an end plate and

Fig . 7 showε a coupling sleeve.

In Figs 3 to 7 each of the particular parts is shown from both axial directions, in order to illustrate certain details.

A hydraulic motor 1 has a shaft 2 which extends over virtually the entire length of a housing 3.

A gear wheel 4, which has a first external toothing 5, is arranged on the shaft 2 so that they rotate together. This first external toothing 5 engages with a first internal toothing 6 which is arranged on the inside of a gear ring 7. The first external toothing 5 and the first internal toothing 6 together form a first toothed coupling.

The gear ring 7 in its turn has a second external toothing 8 which engages with a second internal toothing 9 formed in a toothed rim 10 which in turn forms part of the housing 3. For example, the toothed rim 10 can be arranged as a plate between two housing

parts 11, 12 and be fixed axially by means of screw bolts 13.

The second external toothing 8 and the second internal toothing 9 together form a second toothed coupling. In both toothed couplings the external toothing 5, 8 has fewer teeth than the internal toothing 6, 9.

In the preεent caεe, between the first external toothing 5 and the first internal toothing 6 there are formed pressure pockets 14, the number of which corresponds to the number of teeth of the first external toothing 5. These pressure pockets 14 must now be supplied in the correct order with hydraulic fluid which is supplied and discharged through connections 15, 16. Depending on which direction of rotation of the motor is desired, one of the two connections is used as the input connection and the other is used as output connection.

The principle of operation of such a motor is known. A pressure pocket 14 which is enlarging for kinematic reasons has hydraulic fluid supplied to it under pressure. A presεure pocket 14 that is reducing for kinematic reasons has hydraulic fluid removed from it and discharged through the other connection. Here, the gear ring 7 rotates in the same direction as the gearwheel 4. At the same time, it orbits in the opposite direction. In this manner it is possible to achieve very low speeds combined with a very large torque.

For supplying the pressure pockets 14 in the correct order, that is, for commutation, a piston 17, illustrated in further detail in Fig. 5, is provided in the housing 3. The piston has a central opening 18

through which the shaft passeε. The piεton 17 itεelf is held non-rotatably in the housing 3 by means of a pin 19. Further, the piston 17 has two steps which are sealed in the housing with respect to one another by means of seals 20-22. In each step there is a εerieε of axial ducts 23 to 24 which are arranged in a circle and open into a contact face 25 on the side of the piston 17 remote from the steps. In this case the one connection 15 is in connection by way of an annular channel 26 with the inner boreε 23, whilst the other connection 16 is connected by way of an annular channel

27 to the outer bores 24.

A rotary slide valve 28, which is illustrated in more detail in Fig. 4, lies adjacent to the contact face 25. The rotary slide valve 28 haε two annular grooves 29, 30 on its εide adjacent to the piston 17, each of which is arranged so that it is in register with bores 23, 24 respectively of the piston 17. The rotary slide valve

28 can therefore rotate relative to the piston 17 without interrupting a fluid connection between the annular grooves 29, 30 and the bores 24, 23 in the piston 17.

Further, the rotary slide valve 28 has projections 31 projecting radially outwards, the function of which will be explained later.

Reεpective bores 32, 33 which are arranged at an angle to the axial direction are provided in the annular grooves 29, 30. All bores 32, 33 accordingly open out on the other side of the rotary slide valve 28 in a common circle and there form control openings. On that side of the rotary slide valve 28 (illustrated in Fig. 4b) the control openings 32, 33 are therefore supplied with high and low preεεure alternately.

It is also possible, of course, for one of the two annular grooves 29, 30 to be arranged axially above the control openings 32, 33 so that not all the ducts, but only every other duct, has to run obliquely.

The rotary slide valve 28 lies with its side remote from the piston 17 against a lateral plate 34, which is illustrated in greater detail in Fig. 3. The rotary slide valve 28 lies in this case with its side illustrated in Fig. 4b against the side of the lateral plate 34 which is illustrated in Fig. 3a.

The lateral plate 34 has a number of apertureε 35 correεponding to the number of pressure pockets 14. The lateral plate 34 is connected to the gearwheel 4 by means of screwε 36 so that each aperture 35 opens into a gap in the first external toothing 5. The number of control openings 32 and the number of control openings 33, however, corresponds to the number of teeth of the first internal toothing 6.

On the side of the gear assembly comprising gearwheel 4 and gear ring 7 lying opposite the lateral plate there is arranged a second lateral plate 37 which looks the same as the first lateral plate 34 which is shown in Fig. 3. Further discussion of the plate illustrated in Fig. 3 is therefore sufficient.

The screws 36 pass through bores 38 in the gearwheel 4 and also through bores 39 in the lateral plates 34, 37. The two lateral plates 34, 37 and the gearwheel 4 are thus axially fixedly connected to one another. These parts also rotate together.

On the side remote from the gearwheel 4, illustrated in Fig. 3a, between the apertures 35 there are auxiliary

openings 40 which are connected by way of ducts 41 to the bores 39 for receiving the εcrews 36. Since openings 32, 33 with high and low pressure are arranged alternately on the corresponding end face of the rotary slide valve 28, there is a corresponding distribution also in the apertures 35 and the auxiliary openings 40 respectively. A relatively uniform preεsure distribution over the end face of the lateral plate 34 iε thus achieved.

The corresponding pressures are able, however, to reach through the screw holes 39 in the lateral plates 34, 37 and bores 38 in the gearwheel 4 to the other axial side of the gearwheel 4 aε well, εo that the preεsure on both sides is the same. The preεsureε in the preεsure pockets 14 are also able to reach to the other axial side, because in each presεure pocket there is a presεure compenεation duct 42 in the form of a groove in an inter-tooth space of the external toothing 5.

The lateral plate 37 is covered on the side remote from the gearwheel 4 by an end plate 43, illuεtrated in Fig. 6. The end plate 43 haε a number of slots 44 corresponding to the number of openings 35 and auxiliary openings 40. On the opposite side 45 the end plate 43 is of a spherical shape which lies in a complementary counter-shape 46 in the housing 3. On that side 45 drainage grooves 47 are provided, through which fluid that has leaked at the centre of the end plate 43 can be discharged to the outside.

The spherical shape of the side 45 and the correεponding bearing εurface 46 in the houεing enable the shaft 2 to bend slightly, without the journalling of the entire unit in the houεing 3 leading to canting moments.

In operation, the shaft 2 rotates with the gearwheel 4, whereas the gear ring 7 rotateε and orbitε. The rotary εlide valve 28 iε to rotate together with the gear ring 7 without performing the orbiting movement thereof. For that purpose a coupling sleeve 48 is provided, illuεtrated in greater detail in Fig. 7. In the region of one of its end faces, the coupling sleeve 48 has an inwardly projecting circumferential projection 49, which has two cut-out portions 50. In correspondingly opposing sectionε 51, the second external toothing 8 of the gear ring 7 has teeth of a greater thickness than the remaining teeth. This can be achieved by grinding down the gear ring in the region of its second external toothing 8 somewhat. These teeth in the regions 51 then form projections onto which the cut-out portionε 50 of the coupling εleeve 48 can be placed. The coupling εleeve 48 then follows the rotating and orbiting movement of the gear ring 7. Because the teeth and thus the sectionε, are arranged far towardε the outεide, a relatively large moment can eaεily be transmitted here.

On the axially opposite side, the coupling sleeve 48 has a similar circumferential radially inwardly projecting projection 52, having in its turn recesses 53. The projections 31 of the rotary slide valve project into these recesses 53. One can easily see that with such a coupling, the rotating movement of the gear ring 7 and thus of the coupling sleeve 48, but not the orbiting movement, is transmitted to the rotary slide valve 28.

The shaft 2 iε mounted in the houεing on both sides of the gear assembly, to be precise, on the right-hand εide in Fig. 1 by means of a needle roller bearing 54 and on the left-hand side in Fig. 1 by means of two

roller bearings 55, 56 which are biasεed by a nut 57. This is the output end of the shaft 2.

The motor 12 operateε in known manner. Becauεe the coupling εleeve 48 rotates synchronously with the gear ring 7 and the lateral plate 34 rotates synchronouεly with the gearwheel 4, the preεεure pocketε 14 are supplied in the correct order, that is, commutation is effected. Non-uniform hydraulic pressures have virtually no adverse effects any more. Gearwheel 4 and gear ring 7, together with the two lateral plates 34, 37, form an axially fixedly connected unit, so that no asymmetrical forces are able to develop here. All axial forceε are abεorbed by the screws 36. The unit itself is symmetrically loaded by hydraulic presεureε on both axial εideε. These pressures arise on the one side between the one lateral plate 34 and the rotary slide valve 28 and on the other hand between the other lateral plate 37 and the end plate 43. Tilting moments on the unit are therefore avoided.