EP1074740A1 - Hydrostatic rotary piston machine - Google Patents

Abstract

The machine has a displacement part (1) working with a centrally supported shaft (2). The displacement part has a first rigid casing part (4) with a first inner toothing, which co-operates with a rotatable eccentric rotary piston having an outer toothing with one more tooth than the first inner toothing. The rotary piston has a second inner toothing (8), which meshes with a second outer toothing (9) at the shaft, where the second inner toothing has two more teeth than the second outer toothing. Shaft bearings (10,11) are arranged directly on the left and right of the displacement part. A disc-shaped turning valve and an eccentric inner gear (12) are provided for the drive, where the inner gear guides the rotary valve in orbit about the machine axis of the eccentric movement.

Description

The invention relates to a hydrostatic rotary piston machine according to the preamble of claim 1.

Such a rotary piston machine is for example in the EP-A1-761 968. In this machine, the displacer part and the control section between the shaft bearings for arranged the drive and output shaft penetrating both parts. The advantage of this arrangement is that a Large bearing distance arises, resulting in additional radial forces at the outer end of the shaft e.g. by belt or tooth forces or through wheel contact forces the bearing loads be reduced. Another advantage of this machine is the much better mechanical-hydraulic Starting efficiency compared to other known systems of the so-called orbit slow-speed, mostly with a cardan shaft the torque from the rotary piston to the output shaft transfer.

However, it has been shown that in known machines such as for example according to CH-A5-679062 or EP-A1-0 761 968 a significant increase in pressure and thus performance compared to the other low-speed orbits (high-torque motors) is not possible because the tooth force on the shaft due to the large hydrostatic force of the rotary piston, excessive shaft deflection, bending and shear stress causes. The wave bending leads additionally not evenly distributed over the tooth width Tooth flank pressure, which is why the life of this gear is reduced.

The object of the invention is to improve this machine so that it is higher than the known version Working pressures and therefore higher torques and outputs are possible with a reduced number of components.

This results in reduced manufacturing costs and one very compact design. The aim is a so-called "high torque motor" for a maximum pressure of approx. 400 bar and for a continuous pressure of 350 bar. This demand hangs together with the fact that such hydraulic motors with today's Axial and radial piston pumps must be operated as adjustable hydrostatic power units in many cases be used. That means the machine is essential more robust and at the same time the volumetric Efficiency can be improved.

The design of such an engine does indeed 4 of EP-A1-0 761 968 this demand in a certain Mass already account, but the training is as such relatively expensive, as shown below.

Are the rolling bearings of the hydrostatically radially highly loaded Part of the shaft - as in the training according to Fig.4 EP-A1-0 761 968 shown - immediately adjacent with a small one Axial distance are arranged, the rotary valve must be directly from the shaft can be driven with a gear transmission that allows the rotary valve with the rotary piston of the displacement part turns precisely and synchronously. For commutation the supply and disposal of the work cells of the orbit principle such machines are essential. In this way the shaft deflection and the Misalignment of the shaft tooth flanks reduced under load. Furthermore, the shaft teeth on the displacer part can do the same wide or even a little wider like the rotary piston. This was the case with well-known machines, such as theirs for example, are not described in CH-A5-679062 possible because part of the force-transmitting tooth width there on the shaft due to the meshing of the connecting shaft from the rotary piston to the rotary valve is lost. The Tooth joint bending stress and the specific tooth flank load but can by the measure described by 15 to Be reduced by 20%. Another advantage is the saving the connecting shaft between the rotary piston and the Rotary valve, which accounts for about 3 to 5% of the manufacturing costs.

The invention has now taken on the task of To resolve known disadvantages of the prior art. This is achieved through the realization of the characteristic Features of claim 1

Alternative or advantageous training are through the Features of the dependent claims described.

According to the invention, an eccentric internal gear is available as the gear transmission at which the disc-shaped rotary valve executes the eccentric movement (orbit movement). The two Internal gears that form the eccentric gear have differences in the number of teeth between one and two teeth, making one There is multiple tooth engagement, similar to the positive toothing on the displacer. Cycloid internal gears can be used as tooth shapes, especially trochoid gears, be used, or with a difference in the number of teeth of two teeth also involute splines according to DIN 5480 with 30 ° pressure angle if it is ensured that no tooth head interference disorders occur.

Is the disc-shaped rotary valve by powder metallurgical Sintering process is made for these gears no additional manufacturing effort required. The gearing on the shaft with rational production in one Clamping together with the shaft toothing for the Displacement part on program-controlled gear cutting machines be made. If necessary, this gear can also be punched or sintered onto the shaft to prevent rotation become. The ring gear with the internal toothing in the housing part the rotary valve is punched or cleared, one large number of parts can be cleared at the same time. Here too, the manufacturing effort is minimized.

In order to ensure the same speeds of the rotary piston and rotary valve, the number of teeth on the toothing of the equation should b a · d - c d - c = x w · e.g. - y e.g. - y correspond, where a is the number of teeth of the external teeth on the shaft, b the number of teeth of the internal teeth on the rotary piston, c the number of teeth of the external teeth on the rotary piston, d the number of teeth of the internal teeth on the rigid housing part, w the number of teeth on the first sun gear on the shaft, x the number of teeth the internal teeth on the rotary valve, y is the number of teeth on the external teeth on the rotary valve, and z is the number of teeth on the second sun gear designed as a fixed ring gear. According to the invention, this expression of the equation should be an integer.

In contrast to the known designs of rotary piston machines with disk-shaped rotary valve leads in the inventive Machine, as mentioned above, the rotary valve a little eccentric movement. This may be the proper one Do not negatively influence commutation for the displacer. It should be kept as small as possible. At the pitch circle diameter given by the overall construction for the teeth of the eccentric gear common eccentricity the greater the lower the Speed of the eccentricity axis. The opposite is the case The speed of the eccentric axis is higher, the smaller the eccentricity is chosen. Effect high speeds of the eccentric axis Centrifugal forces on the rotary valve, creating the gears are loaded and crushing losses between the internal gears arise. Preferred gear data are by given the characteristic of claim 4.

A good compromise between the size of the eccentricity and the speed of the eccentric axis is through the realization the characterizing features of claims 5 to 7 possible. With these values, the centrifugal forces are on the Rotary valve still small, but at the same time in a large one Angular range of the rotary piston the commutation conditions improved for a low-torque pulsation run the machine and good volumetric efficiency.

The stepped piston hydraulically compensates for the axial forces on disc-shaped rotary valve, so that in both directions of rotation the leakage gap between the rotary valve and the control plate on the one hand and the rotary valve and the end face the stepped piston on the other hand to a lubricating film thickness of is reduced by a few micrometers. That way stays the volumetric efficiency of the machine even at high Pressures and low speeds very high.

Fig.1

einen Längsschnitt durch eine erfindungsgemässe Kreiskolbenmaschine,

Fig.2

einen Querschnitt längs A-A der Fig.3,

Fig.3

eine Teilansicht entsprechend Fig.1, Blickrichtung von oben,

Fig.4

eine schematische Darstellung eines erfindungsgemässen Drehventils;

Fig.5

ein Längsschnitt entsprechend Fig.1 einer weiteren Ausbildungsvariante und

Fig.6

einen Teil-Längsschnitt durch eine weiter Ausbildungsvariante

The invention is described in the following purely by way of example with reference to drawings. Show it:

Fig. 1

2 shows a longitudinal section through a rotary piston machine according to the invention,

Fig. 2

3 shows a cross section along AA of FIG. 3,

Fig. 3

a partial view corresponding to Figure 1, viewing direction from above,

Fig. 4

a schematic representation of a rotary valve according to the invention;

Fig. 5

a longitudinal section corresponding to Figure 1 a further training variant and

Fig. 6

a partial longitudinal section through a further training variant

The rotary piston machine shown in the figures has an input or output shaft 2, in which the bearing 10 directly are arranged on both sides of a rigid housing part 4. The Shaft 2 is in the area of the rigid housing part 4 with a - second - external toothing 9 with a number of teeth a, with a - second - Internal teeth 8 with a number of teeth b on the rotary piston 6 combs. The rotary piston 6 circles eccentrically around the shaft 2 and meshes with a - first - external toothing 7 with a Number of teeth c in the - first - internal toothing 6 with one Number of teeth d of the rigid housing part 4.

To increase the service life of the positive gear between Rotary piston 6 and housing part 4 can - the first - internal teeth 5 on the housing part 4 in an advantageous manner Be designed in the form of rotatably sliding rollers 28.

Allow the unloaded rollers on the high pressure side the construction of a lubricating film between roller 28 and Housing4, the one with intermittent relief of the roller 28 tragic squeeze film produced (squeeze effect). Consequently there is a real hydrodynamic plain bearing.

A gear transmission is used to transmit the rotary piston rotation provided in the form of an eccentric gear 12, 13, whereby a translation is generated via this gear transmission is through which the translation in the rotation transmission is compensated by the rotary piston 6 on the shaft 2. The such driven control part in the form of a rotary valve 3 is disc-shaped. Comparing the invention Training according to the present Fig.1 and FIG. 5 with that known from EP-A1-0 761 968, FIG. 5 Training, it is easy to see that the inventive Training with fewer components (so could on the transmission piston 15 ', the transmission sleeve 42 and also the housing part 9 of the known training can be dispensed with) is, thereby reducing the manufacturing cost. The size of the motor can also be chosen to be smaller become.

4 schematically shows the design and mode of operation of the rotary valve 3. The shaft 2 is designed as the first - sun gear 14 with a number of teeth w, into which the disk-shaped rotary valve has an - third - internal toothing 15 with a number of teeth x with an eccentricity 20 engages in rotation. The rotary valve 3 now meshes with its - third - external toothing 16 with a number of teeth y with a - fourth - internal toothing 17, which is formed in the connection housing 18 on a second sun gear 18 ¹ in the form of a ring gear fixed to the housing. The number of teeth of this - fourth - internal toothing 17 on the second sun gear 18 ¹ is given by z. As shown further below, the internal toothing provided on the second sun gear can also be arranged directly on the connection housing, as a result of which the sun gear as a separate component, although fixed to the housing, can be dispensed with.

When the machine runs clockwise, oil is injected under high pressure into the Port 27 and thus introduced into the annular space 52. Because of the conical design of the rotary valve 3 is one annular axial surface 30 (Fig.5) available in this operating state of the machine, the axial force at the leakage gap 24 can compensate. The rotary valve 3 thus remains always at zero clearance at the leakage gap 24 when the axial surface 30 is correctly dimensioned.

It is different if it is under high pressure when turning to the left Oil is added to port 26 (Fig.1 and Fig.3). In this case, the pressure oil acts on a step piston 23 formed piston ring surface 32. This stepped piston 23 is for axial compensation of the leak gaps 24 and 25 provided on the rotary valve 3. Since the stepped piston 23 is axially movable sealed in the connection housing 18 by 0-rings 33 and 34 is, this is pressed against the leakage gap 25 to Sealing another, assigned to the piston ring surface Annulus 35. The axial force generated by the stepped piston 23 thus seals the leak gap 24 and the leak gap at the same time 25. Function and dimensioning of such a stepped piston 23 is known to the person skilled in the art and therefore needs not to be explained in more detail.

5 is a new version of the step piston 23 shown. The stepped piston 23 must be prevented from rotating be secured in the connection housing 18. It serves according to the design of Figure 1 a pin 53. However can then no connection bore 36 at this point Step piston 23 are attached, which is the flow resistance enlarged for the oil. The extent of the Pin 35 against the connector housing 18 by the friction torque between the rotating rotary valve 3 and the Step piston 23 is caused, the higher the higher the Working pressure of the machine is given the axial force on the Piston ring surface 32 increases. This scope hampers the axial mobility of the stepped piston 23 in disadvantageous Wise.

In Figure 5, the ring gear-shaped, second sun gear 18 ^{1 is made} slightly wider. In this way, a toothing 37 provided on the stepped piston 23 can engage in the stepped piston 23, so that the stepped piston 23 can be prevented from rotating against lateral forces. The stepped piston 23 can be manufactured together with its toothing 37 and the connecting bores 36 in the sintering process. This version is very easy to install, since the stepped piston 23 can be used if the 0-rings 33 and 34 are arranged in the connection housing 18. The anti-rotation device for the second sun gear 18 ¹ is done via pins 38 and its axial securing device via a snap ring 39. An initial spring 40 is designed as a corrugated spring and keeps the stepped piston 23 in contact with the rotary valve 3 even at zero pressure.

The fourth internal toothing 17 can also be arranged directly in the connection housing 18, as shown in FIG. This will prove to be particularly advantageous from the point of view if the manufacturer of such circular piston machines has sufficient capacity for gear-butting machines. Such an arrangement has the advantage that the internally toothed sun gear 18 ¹ and small parts such as pins 38 and locking ring 39 can be saved. This also reduces the effort involved in assembly. In the design according to FIG. 6, it should be noted that the relative twist phase position of the toothing 17 with respect to the second internal toothing 5 (which can be designed in the form of rollers according to FIG. 2) is exactly observed, which task in the training according to FIG. 1 and 5 is fulfilled by the correct positioning of the pin 38.

The slight eccentric movement of the rotary valve 3 has a tribological effect very advantageous because it causes scoring Dirt and abrasion particles in the oil film is avoided, such as when polishing smooth surfaces. The inevitable Wear due to erosion and corrosion on these surfaces is by the hydrostatically pressed stepped piston 23 automatically readjusted. This keeps the leakage current always small at these points.

Claims (9)

Hydrostatic rotary piston machine with a Output part acting displacement part (1) and with a for supplying and disposing of the displacement part (1) with Working fluid serving rotary valve (3), the displacement part (1) a first rigid housing part (4) with a has first internal toothing (5) with a number of teeth d, those with a first tooth number c External teeth (7) on a rotatable, eccentric arranged rotary piston (6) interacts, the Orbital piston (6) has a second internal toothing (8) with a Has number of teeth b with a second, one number of teeth a having external teeth (9) on a centric mounted shaft (2) and where d - c = 1 and b - a = 2, with shaft bearings (10, 11) being immediate arranged on the left and right of the displacer (1) are and a disc-shaped rotary valve (3) and a gear transmission is provided to drive it, thereby characterized that the gear transmission is an eccentric internal gear (12, 13), in which the disc-shaped Rotary valve (3) in orbit around the machine axis the eccentric movement executes. Hydrostatic rotary piston machine according to claim 1, characterized in that the number of teeth (a, b, c, d) of the displacement part (1) and the number of teeth (w, x, y, z) of the eccentric internal gear (12, 13) satisfy the equation b a · d - c d - c = x w · e.g. - y e.g. - y and this expression is a whole positive number, where w is the number of teeth of a first sun gear (14) arranged on the shaft (2), x the number of teeth of the third internal toothing (15) on the disk-shaped rotary valve (3), y the number of teeth of the third External teeth (16) on the rotary valve (3) and z means the number of teeth of the fourth internal teeth (17) on the connection housing (18). Hydrostatic rotary piston machine according to claim 2, characterized characterized that the positive integer is equal 3 is. Hydrostatic rotary piston machine according to one of claims 1 to 3, characterized in that the value of the equation y y - e.g. Integer negative values between -33 and -55 assume when y on the eccentric gear (12, 13) y the number of teeth on the external toothing (16) on the disk-shaped rotary valve (3) and z the number of teeth on the internal toothing (17) on the connection housing (18) or one mean second sun gear (18 ¹ ) in the connection housing (18). Hydrostatic rotary piston machine according to one of claims 2 to 4, characterized in that at the number of teeth a = 12, b = 14, c = 11, d = 12 or a = 13, b = 15, c = 12, d = 13 of the displacement part (1) the number of teeth of the eccentric gear (12, 13) of the rotary valve (3) can assume the following values:
at (x - w) = 1: x = 16 to 24 y = 29 to 45 teeth
at (x - w) = 2: x = 31 to 49; y = 18 to 46 teeth. Hydrostatic rotary piston machine according to one of the claims 1 to 5, characterized in that the common Eccentricity (20) of the eccentric gear (12, 13) on disc-shaped rotary valve (3) 0.01 to 0.017 times the mean pitch circle diameter of control slots (21) in a control plate (22). Hydrostatic rotary piston machine according to one of the claims 1 to 5, characterized in that the common Eccentricity (20) of the two internal gears (12, 13) 0.011 to 0.015 times the mean pitch circle diameter the control slots (21) in the control plate (22). Hydrostatic rotary piston machine according to one of the claims 1 to 7, characterized in that a stepped piston (23) is provided for axial compensation of leakage gaps (24 and 25) on the disc-shaped rotary valve (3). Hydrostatic rotary piston machine according to claims 1 to 8, characterized in that the first internal toothing (5) by rollers rotatably mounted in the housing part (4) (28) is formed.

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