EP1776525B1 - Hydrostatic rotary cylinder engine - Google Patents

Description

The invention relates to a hydrostatic, low-speed rotary piston engine according to the preamble of the independent claims 1 and 2.

A hydrostatic rotary piston engine of this type is known from EP 1 074 740 B1 known. An advantage of the disclosed there formation of a rotary piston engine is compared to previous solutions is that the bearings of the hydrostatically highly loaded part of the shaft are arranged immediately adjacent with small axial distance in the fixed housing, so a small amount of bending and tooth deformation on the shaft and, accordingly, a maximum Printing performance and thus to torque delivery can be achieved. Because of this bearing arrangement, there is no way to create a 1: 1 rotary connection between the acting as a rotor rotary piston and responsible for the commutation rotary valve, it has been proposed to synchronously drive the rotary valve via a gear transmission from the shaft. In the known embodiment, this gear transmission is an eccentric internal gear, in which the disk-shaped rotary valve itself acts as an eccentric member of this transmission and thus performs an unavoidable orbital movement. Extensive tests have shown, however, that this thought, which at first seems striking, can not be realized in practice at high working pressures, because the necessary eccentric movement of the rotary valve with respect to the fixed control plate does not enable a sufficiently accurate commutation of the machine. The result is highly fluctuating torque output on the shaft, unsatisfactory volumetric efficiency and high noise, since the outer part of the eccentric gear must work in the high pressure range. Also the axial compensation of the on Rotary valve axially acting hydraulic forces through the balance piston was not optimal due to the eccentric movement of the rotary valve.

Since the teeth of the eccentric drive generate a displacement effect, similar to an internal gear pump, it is unfavorable because of the resulting hydrostatic losses when this displacement occurs in the high pressure part of the machine.

The invention has as its object to eliminate these deficiencies and at the same time reduce the caused by the orbital movement slightly increased friction at the rotary valve and the production costs.

This object is achieved by the realization of the characterizing features of the independent claims. Features which further develop the invention in an alternative or advantageous manner can be found in the dependent claims.

The invention eliminates these disadvantages while retaining the above-mentioned advantages of such machines.

The hydrostatic, low-speed rotary piston engine according to the invention comprises a power unit acting as an output with a centric fixed stator, a rotary piston as the rotor and a centrically mounted shaft. The stator has an internal toothing with the number of teeth d. The rotary piston has a part engaging in the internal toothing of the stator outer teeth with a number of teeth c and an internal toothing with a number of teeth b. The shaft meshes with its external teeth with a number of teeth a partially the internal toothing of the rotary piston, wherein the rotary piston for performing an orbital movement is arranged and dimensioned eccentrically such that with working fluid ver and disposable tooth chambers between the inner teeth of the stator and the outer teeth of the rotary piston form. An inlet and outlet part is used for supply and disposal of the power unit with the working fluid. By means of a disc-shaped rotary valve, which according to the invention is mounted centrically running to the shaft and the stator, the control of the supply and disposal of the tooth chambers with the working fluid. In addition, the rotary engine comprises a gear transmission, which is arranged between a shaft outer shaft of a shaft formed by a sun gear with a number of teeth w and an internal toothing of a fixed internal gear with a number of teeth z as a synchronous drive for the rotary valve. The shaft is mounted on both sides of the power section immediately adjacent bearings arranged. According to the invention, the gear transmission is arranged exclusively in the leakage oil region of the engine and is arranged by a planetary gear with at least one planet carrier, which is rotatably connected to the rotary valve and on which planet gears between the shaft outer teeth and the stationary inner ring gear are arranged, or preferably by an eccentric with an eccentric , which is rotatably connected to the rotary valve formed.

Since a continuous shaft with large shaft diameters and high torsional strength can be used in the hydrostatic, low-speed rotary piston engine according to the invention, it is possible to expose both shaft ends to high torque flow and, for example, both shaft ends as output, or one shaft end as output and the other shaft end for connection to one Brake or a second drive, whereby the entire drive unit can be made considerably more compact.

Because of the allowable by the invention omission of the orbit movement of the rotary valve, by the placement of the eccentric in Leckölraum of the engine and by the use of cost-effective extruded or sintered parts as gear members thus creates an optimal, compact and inexpensive construction. The drive of the rotary valve 1: 1 to the rotary piston of the power unit via a wobbling cardan-like wave is known from the earlier designs. There, however, the wobble shaft must compensate for the full eccentricity of the rotary piston in the power section, so that a very large wobble angle is created. The inventive wobble gear requires a much smaller eccentricity, which according to the invention is independent of the eccentricity of the rotary piston in the power section, so that this wobble angle is substantially smaller than half of that wobble angle of the earlier construction. Thus, caused by the tumbling and necessarily enlarged gear plays the transmission can be drastically reduced. The resulting rattling noises and the wear are much smaller in the inventive construction.

When using a Exzentergetriebes the eccentric has an internal toothing with a number of teeth x and an external toothing with a number of teeth y and is disposed between the shaft outer teeth and the internal teeth of the fixed internal gear, the eccentric with its internal teeth with the shaft outer teeth of the shaft and with its external teeth with the Internal toothing of the fixed internal gear ring meshes.

Preferably, the eccentric gear as a wobble gear and the eccentric disk-shaped eccentric are formed, wherein the disc-shaped eccentric is rotatably connected via a cup-shaped connecting part with the rotary valve via Mitnehmerverzahnungen in the speed ratio 1: 1.

The following equation represents the speed ratio shaft to the rotary piston or shaft to the rotary valve: $$\frac{\frac{b}{a}\cdot d-c}{d-c}=\frac{\frac{x}{w}\cdot z-y}{z-y}$$

As one can easily see from this equation, the numbers of teeth of the eccentric gear can be carried out quite differently.

A first option, for example, would be the design exactly as for the power section with w = 12, x = 14, y = 11 and z = 12. It only needs to be considered that the eccentricity of the two internal gears are exactly the same. The equation expression is a positive integer, preferably equal to 3. It must also be striven that in this area, the diameter of the shaft is sufficiently large, so that their torsional strength for a possibly connected holding brake for the maximum torque is still sufficient. Here, however, the eccentricity of the transmission is relatively large, so that the wobble angle is correspondingly large. However, then the speed of eccentricity would be quite small.

wobei dieses Verhältnis bevorzugt zwischen -3 und -9 liegt.The relationship between the rotational speed Ne of the eccentricity of the eccentric gear and the rotational speed Nw of the shaft is given by the equation $$\frac{\mathit{ne}}{\mathit{Nw}}=-\frac{w\cdot y}{x\cdot z-w\cdot y}.$$

this ratio is preferably between -3 and -9.

A second option is the preferred design of the number of teeth according to a = 12, b = 14, c = 11, d = 12, w = 12, x = 13, y = 23 and z = 24 or after a = 12, b = 14 , c = 11, d = 12, w = 9, x = 10, y = 17 and z = 18, each with a very small eccentricity. As can be easily calculated from the above equation Ne / Nw, then the rotational speed of the eccentricity becomes higher, but still remains below the value of the wobble wave of earlier known constructions.

In the design of the eccentric with the numbers of teeth a = 12, b = 14, c = 11, d = 12, w = 12, x = 13, y = 23 and z = 24, there are advantages: As in the assembly of the engine Rotary position of the rotary valve must always exactly match the rotational position of the motor power unit in the phase position, it makes sense that the number of teeth w and their rotational position on the shaft is exactly the same as the number of teeth a of the external teeth on the shaft on the power unit and its rotational position. Thus, the shaft can be mounted at all times, without having to pay attention to what rotational position it is, whereby the assembly is considerably simplified.

The proposed numbers of teeth a = 12, b = 14, c = 11, d = 12, w = 9, x = 10, y = 17 and z = 18 have the advantage with respect to the toothing for the eccentric gear that the toothing module becomes larger, the stability of the shaft in this area increases and in particular the negative-running speed of the eccentric axis of the eccentric disc drops sharply, resulting in a smoother running of the transmission. It is thereby accepted that the wobble angle is about larger, and dispenses with the above-described advantage during assembly.

Experiments have shown that very good results are achieved when the common eccentricity of the eccentric gear is 0.013 to 0.015 times or 0.015 to 0.022 times the mean pitch circle diameter of the control slots in the control plate.

Since in the conventional machines with carding shaft between the rotary piston and the output shaft (of which currently about 1.2 million pieces are produced worldwide), the large hydrostatic radial force on the rotary piston completely absorbed by the teeth between the acting as a rotor rotary piston and the stator must be, the Hertzian pressure and thus the friction between these teeth is very large, because the propeller shaft is known to absorb any radial forces. Especially at low speed and high working pressure therefore the friction losses and the wear of the teeth are extremely large. Therefore, the starting efficiency of these machines is correspondingly poor and is only about 63 to 71%.

For high working pressures - especially over 120 bar - it is therefore indispensable in these earlier designs with cardan shaft as a torque connection between the rotary piston and the output shaft, that the teeth of the internal teeth on the stator are formed by rollers in their precisely machined caverns in the stator by a transient hydrodynamic oil film are rotatably mounted. The rollers must be designed with high hardness and best surface quality, as well as the necessary precise caverns in the stator.

In the machine according to the invention, the radial load of the teeth between the rotary piston and the stator is only a fraction of the ratios described above, so that the pressure of the engine can be increased considerably even without roles in the stator. Nevertheless, it is also in the machine according to the invention advantageous if the usual roles in the stator are maintained, which leads to further increased printing performance and excellent life. Measurements have shown that in the machine according to the invention the starting efficiency and also the mechanical-hydraulic efficiency can be increased by 3 to 5% through the transition to rollers in the stator. The start-up efficiency reaches values of more than 90%.

When using the hydrostatic, low-speed high-torque motor according to the invention as a wheel motor, the output-side roller bearing requires a higher radial load capacity for additional reception of the wheel load. It should be located as close to the center of the wheel. Since, for example, in material handling equipment shock-like elevations of the static wheel load can occur, it is advantageous if this bearing is as close as possible to the wheel flange and is optionally arranged outside the leakage space of the rotary piston engine with a permanent rolling bearing grease directly in the housing part of the rotary piston engine.

The rotary piston engine according to the invention is due to the advantageous bearing arrangement and the powerful continuous shaft, inter alia, excellent as a wheel motor or winch drive for direct driving a wheel or a cable drum. In this case, the shaft is preferably formed integrally with a wheel flange on which a wheel or a cable trench for direct drive can be mounted directly.

The device according to the invention is described in more detail below purely by way of example with reference to concrete exemplary embodiments shown schematically in the figures, with further advantages of the invention also being discussed.

Fig. 1

ein erstes Ausführungsbeispiel eines Kreiskolbenmotors mit einem Exzentergetriebe in einem Längsschnitt entlang der Schnittlinie C-C der Fig. 2,

Fig. 1.1

eines zweiten Ausführungsbeispiels eines Kreiskolbenmotors mit einem Planetengetriebe in einen Teil-Längsschnitt entlang der Schnittlinie C-C der Fig. 2,

Fig. 1.2

eine Querschnitt durch das Exzentergetriebe des ersten Ausführungsbeispiels des Kreiskolbenmotors,

Fig. 2

einen Querschnitt entlang der Schnittlinie D-D der Fig. 1 durch das Rotor-Stator-System des ersten Ausführungsbeispiels,

Fig. 3

einen Querschnitt durch das Rotor-Stator-System eines Ausführungsbeispiels mit drehend gelagerten Rollen als Innenverzahnung im Stator,

Fig. 4

eine Ansicht X auf Fig. 1 auf einen SAE-Anschluss eines Ausführungsbeispiels, einen Teilschnitt entlang der Linie A und einen Teilschnitt entlang der Linie B der Fig. 3,

Fig. 5

einen Längsschnitt durch ein Ausführungsbeispiel eines erfindungsgemässen Radmotors,

Fig.6

einen Längsschnitt durch einen Radmotor gemäss der Erfindung mit auf der Welle angekoppelter Parkbremse als Lamellenbremse,

Fig.7

einen Längsschnitt durch einen Radmotor gemäss der Erfindung mit an der Welle angekoppeltem zweiten Motor als 2/3-Stufenmotor,

Fig. 8

einen Querschnitt des 2/3-Stufenmotors entlang der Schnittlinie E-E der Fig. 7,

Fig. 9

ein mögliches Hydraulik-Schaltbild zur Steuerung des 2/3-Stufenmotors gemäss Fig. 7 und Fig. 8 mit beispielhaften technischen Angaben,

Fig. 10

einen Längsschnitt durch einen erfindungsgemässen Kreiskolbenmotor mit einer an der Welle angekoppelten, gross dimensionierten Arbeitsbremse als Lamellenbremse,

Fig. 11

einen Längsschnitt durch eine vorteilhafte Weiterbildung eines erfindungsgemässen Kreiskolbenmotors mit einer umlaufende axiale Entlastungsnut an der axialen Gleitfläche zwischen Drehventil und Ausgleichskolben,

Fig. 12

eine Querschnittsansicht auf den Steuerspiegel der Steuerplatte des Kreiskolbenmotors aus Fig. 11,

Fig. 13

einen Längsschnitt durch das Drehventil und den Ausgleichskolben des Kreiskolbenmotors aus Fig. 11 in einer Detailansicht und

Fig. 14

eine Linksansicht auf das Drehventil und den Ausgleichskolben aus Fig. 13.

In detail show:

Fig. 1

a first embodiment of a rotary engine with an eccentric gear in a longitudinal section along the section line CC of Fig. 2 .

Fig. 1.1

of a second embodiment of a rotary piston engine with a planetary gear in a partial longitudinal section along the section line CC of Fig. 2 .

Fig. 1.2

a cross section through the eccentric gear of the first embodiment of the rotary engine,

Fig. 2

a cross section along the section line DD of Fig. 1 by the rotor-stator system of the first embodiment,

Fig. 3

a cross-section through the rotor-stator system of an embodiment with rotationally mounted rollers as internal teeth in the stator,

Fig. 4

a view X on Fig. 1 to an SAE connection of an embodiment, a partial section along the line A and a partial section along the line B of Fig. 3 .

Fig. 5

a longitudinal section through an embodiment of an inventive wheel motor,

Figure 6

a longitudinal section through a wheel motor according to the invention with coupled on the shaft parking brake as a multi-disc brake,

Figure 7

a longitudinal section through a wheel motor according to the invention with coupled to the shaft second motor as a 2/3 stage motor,

Fig. 8

a cross section of the 2/3-stage motor along the section line EE of Fig. 7 .

Fig. 9

a possible hydraulic circuit diagram for controlling the 2/3-stage motor according Fig. 7 and Fig. 8 with exemplary technical data,

Fig. 10

a longitudinal section through an inventive rotary piston engine with a coupled to the shaft, large-dimensioned working brake as a multi-disc brake,

Fig. 11

a longitudinal section through an advantageous development of an inventive rotary piston engine with a circumferential axial relief groove on the axial sliding surface between rotary valve and balance piston,

Fig. 12

a cross-sectional view of the control mirror of the control plate of the rotary piston engine Fig. 11 .

Fig. 13

a longitudinal section through the rotary valve and the balance piston of the rotary piston engine Fig. 11 in a detailed view and

Fig. 14

a left view of the rotary valve and the balance piston Fig. 13 ,

In the following, possible embodiments with reference to several figures, which partially show a single embodiment in different views with sporadically different levels of detail, explained, reference being made in part to already mentioned in previous figures reference numerals.

Fig. 1 shows a first embodiment of an inventive rotary piston engine with an eccentric gear in a longitudinal section, while Fig. 2 a cross section through the rotor-stator system of the first embodiment along the section line DD of Fig. 1 shows. Furthermore is in Fig. 2 the cutting direction of the Fig. 1 seen from the section line CC in. The rotor-stator system of the power unit 1 of the rotary piston engine comprises a centric fixed stator 4 with an internal toothing 5, hereinafter referred to as first internal toothing 5, in which an eccentrically arranged for performing an orbit movement, acting as a rotor rotary piston 6 with a hereinafter as first external toothing 7 called external teeth at least partially engages. A centrally mounted by means of two on both sides of the power unit 1 immediately adjacent bearings 10, 11 mounted shaft 2 has an external toothing 9 - the second external teeth 9 -, which in turn at least partially engages in an internal toothing 8 of the rotary piston 6, called the second internal toothing 8. The forward direction of rotation of the rotor-stator system of the rotary piston engine is defined for the following explanations as the direction of rotation in which the rotary piston 6 in the direction of rotation 60 and the shaft 2 in the direction of rotation 61 in accordance Fig. 2 rotate. Accordingly lie in the Fig. 2 the expanding swallow cells between the first internal toothing 5 and the first external toothing 7 Always on the left and the compressing feed cells always to the right of an eccentric axis 62. Since the eccentric axis 62 performs one of the direction of rotation 61 of the shaft 2 and the direction of rotation 60 of the rotary piston 6 entgegensetze direction of rotation 64, creates a rotating field for the radial hydraulic force on the rotary piston 6, if always High pressure is supplied to the expanding swallow cells. The control of this rotating field worried a rotary valve 3 as a commutator, similar to a DC motor. To cause a forward rotation of a fluid - in particular pressure oil as working fluid - a high pressure port 55 in an inlet and outlet part 70 and thus a first annular space 56 is supplied, which surrounds the rotary valve 3 sealed. According to the numbers of teeth of the first internal toothing 5 of the stator 4 and the first external toothing 7 of the rotary piston 6 in the first embodiment, the rotary valve 3 has eleven circumferentially uniformly distributed, with the first annulus 56 in communication high-pressure window 21a.

A control plate 22 with control slots 21 has twelve uniformly distributed on the circumference pressure window 33 a, which are connected via feed bores 33 with the twelve toothed chambers between the first internal toothing 5 of the stator 4. Because of the circumferential distribution eleven to twelve of the high-pressure window 21a of the rotary valve 3 and the pressure window 33a of the control plate 22 is always only one half of the tooth chambers of the stator 4 under high pressure, and that with the correct phase position of the rotary valve 3 with the rotary piston 6 always those tooth chambers, in of the Fig. 2 lie to the left of the eccentric axis 62. Since the rotary valve 3 between the high-pressure windows 21a uniformly distributed identically designed low-pressure window 21b has, the other half of the twelve tooth chambers of the stator 4 via connecting holes 58a having a ring grooves 108 and 109 second Annulus 58 and thus connected to a low pressure port 57, so that the compressing feed cells displace the fluid under low pressure in the low pressure side and thus in the low pressure port 57.

It should therefore be ensured that the separation axis of the rotary valve 3 in a high-pressure side and a low-pressure side as exactly as possible the same speed and direction of rotation as the rotor-stator system. This requirement is met if the rotary valve has the same direction of rotation and the same speed as the rotary piston 6 about its own axis. In the case of the rotary piston engine according to the invention, in a preferred embodiment, the shaft 2 is mounted roller-mounted directly in the housing on the left and right of the rotor-stator system, so that the drive of the rotary valve 3 must take place via the shaft 2, which due to the system performs a different rotational speed than the rotary piston 6. In the illustrated embodiment, the shaft 2 runs three times as fast about its axis as the rotary piston 6 about its own axis. Accordingly, the rotary engine according to the invention requires a transmission between the shaft 2 and the rotary valve 3 with the same ratio to the slow. This can be done by means of an eccentric gear 30, as in the first embodiment Fig. 1 and Fig. 1.2 , or by means of a planetary gear 80, as in a second embodiment according to Fig. 1.1 shown, happened.

Fig. 1.1 shows the second embodiment of an inventive rotary piston engine with a planetary gear 80 in a partial longitudinal section along the section line CC of Fig. 2 , The planetary gear 80 includes a sun gear 13 on the shaft 2, the shaft outer teeth 14 meshing with planetary gears 90 which are mounted on a planetary carrier 91, which rotates 1: 1 with the rotary valve 3 is coupled. The planet gears 90 mesh simultaneously with a fixed internal gear ring 92, which has twice the number of teeth as the sun gear 13 on the shaft 2. According to the laws of the planetary gear is then the translation of the shaft 2 to the rotary valve 3 exactly 3: 1 slow.

However, it is preferred, as in the first embodiment in Fig. 1 and Fig. 1.2 shown, a simpler constructed eccentric gear 30 is used, which includes a sun gear 13 on the shaft 2 with a shaft outer teeth 14 and a fixed inner ring gear 28, the inner teeth 17, hereinafter called fourth internal teeth 17, compared to the number of teeth of the shaft outer teeth 14 double the number of teeth. Interposed is a disk-shaped eccentric 26, which has an internal toothing 15 in the interior - the third internal toothing 15 - and outside an external toothing 16, referred to as the third external toothing 16, has. Preferably, this eccentric gear 30 is executed with tooth shapes that allow the number of teeth difference between the shaft outer toothing 14 and the third inner toothing 15 and the third outer toothing 16 and the fourth inner toothing 17 is equal to 1. With involute teeth such transmissions are usually not feasible, since in this case, take place tooth interference disorders. Also, they do not allow exact radial centering of the wheels against each other under these conditions. It should therefore be resorted to other tooth shapes. In the example of Fig. 1.2 It is preferred to use a double cycloidal internal external toothing, as described for example in German Pat DE 39 38 346 is known, to which reference is hereby made.

This eccentric gear 30 also has a reduction between the shaft 2 and the disc-shaped eccentric 26 of exactly 3: 1 slow. How out Fig. 1 recognizable, the disc-shaped eccentric 26 is 1: 1 rotationally connected via a cup-shaped connecting part 27 rotatably connected to the rotary valve 3, wherein Mitnehmerverzahnungen 31 and 32 allow the cup-shaped connecting part 27 together with the disc-shaped eccentric 26 a small wobbling movement corresponding to the eccentric movement of the disc-shaped eccentric 26th performs. The backlash of the shaft outer teeth 14, the third internal teeth 15 of the eccentric 26, the third external teeth 16 of the eccentric 26, the fourth internal teeth 17 of the internal ring gear 28 and the Mitnehmerverzahnungen 31 and 32 are designed to be somewhat larger than usual because of the wobbling motion.

In order for the rotary valve 3 to be rotatably but well sealed axially against leakage from the high pressure, an axial compensating piston 65 is provided in a known manner.

In Fig. 3 shows a cross section through the rotor-stator system of another embodiment, in which rotatably mounted rollers 81 are used as the first internal toothing 5 in the stator 4. These rollers 81 should always be trapped in their cavities 82 in the stator 4, ie the caverns 82 should extend in the direction of the shaft 2 beyond the roller radius, so that the rollers 81 do not move radially inward out of the caverns 82 can. This would lead to a blockage of the rotary engine. In Fig. 3 the shape of the caverns 82 is clearly illustrated.

How to get out of the Fig. 2 and 3 can recognize, in a compact construction of the inventive Rotary piston engine with correspondingly small

Be the pitch circle diameter of the screws, the first internal toothing 5 of the stator 4 in the transition to rollers 81 as teeth in the stator 4 are offset by half a pitch, as in Fig. 3 seen. This means that the feed bores 33 and the associated pressure windows 33a and control slots 21 are correspondingly offset on a pitch circle in the control plate 22. Therefore, it is advantageous if the number of teeth of Mitnehmverzahnungen 31, 32 is twice as large as the number of teeth c of the first outer toothing 7 of the rotary piston 6 of the power section 1. In this interpretation, the number of teeth of Mitnehmverzahnungen 31, 32 can then in all cases the rotary valve and the control plate 22 can be used without change. In the case of the preferred design with the numbers of teeth a = 12, b = 14, c = 11, d = 12, w = 12, x = 13, y = 23 and z = 24 or a = 12, b = 14, c = 11, d = 12, w = 9, x = 10, y = 17 and z = 18 then the number of teeth of Mitnehmverzahnungen 31, 32 with 22 to choose.

The housing parts, which include a bearing flange 25, the stator 4 and the input and Auslassteil 70 must be centered against each other during assembly. In Fig. 3 and in Fig. 4 showing a view X on an SAE connection, a partial section along the line A and a partial section along the line B of FIG Fig. 3 In addition, it is shown that two out of the twelve screws are designed as fitting screws 93 which are to be used first during assembly of the motor. From the Fig. 4 is also in partial section A of Fig. 3 It can be seen that the rotary piston engine should be designed very compact due to the specified by the international SAE standard hole patterns for mounting the engine so that dimensions and weight are optimized. A flange screw connection for the high and low pressure connection 55 or 57 according to SAE standard is also shown here.

An application for the rotary piston engine according to the invention is the use as a wheel motor, as in its simplest form as a longitudinal section in Fig. 5 is shown. Extremely advantageous in this embodiment of a wheel motor is the formation of a driven-side roller bearing 11 outside a leakage space 85 directly in the housing part 84 of the engine. Since such wheel motors do not require high speeds, a permanent rolling bearing fat filling is sufficient as lubrication, which is sealed by a NILOS ring 72 to the outside. By this construction, it is possible that a wheel flange 40 can be made integral with the shaft 2, so that for large wheel loads, the shaft 2 is very robust ausbildbar.

In the case of a wheel motor according to Fig. 5 is usually a right- and a left-handed version required. Again, it is advantageous if the rotary valve can be offset during assembly by half a pitch, so that hereby with the same pressure connection and thus the same flow direction of the working fluid, the direction of rotation of the motor is reversible for the same physical operating conditions.

A hydrostatic wheel bearing usually requires a spring-loaded, automatically spring-loaded parking brake, which is independent of the hydraulic pressure, in order to prevent the parked vehicle from rolling away. The Fig. 6 shows a possible realization of such a wheel motor in longitudinal section, in which on the side opposite the output a spring-loaded parking brake 42 is arranged in the form of a multi-disc brake. The rotary piston engine according to the invention advantageously enables a continuous shaft 2 suitable for high torques with a large-dimensioned wool extension 41, so that the lamellae of the parking brake 42 directly via a hub 73 can transmit their braking torque to the shaft 2. In this case, the shaft outer toothing 14 for the eccentric gear 30 is extended outwardly in manufacturing technology, on which the hub 73 can be wedged torque-effective torsionally effective. This spring-loaded parking brake 42 is a wet-running multi-disc brake, which can be released with greatly reduced hydraulic pressure via the separate port 43. As a spring, a plate spring 74 is provided here. How to get out Fig. 5 and 6 can recognize, the fixed fourth internal teeth 17 for the eccentric gear 30 is incorporated directly into the input and output part 70, for example by means of a gear-impact machine or by means of a broach. This results in the advantage that the shaft outer teeth 14 on the shaft 2 in diameter is larger, so that the shaft extension 41 receives a larger torque capacity. This is of particular importance, in particular with wide moving sets in the power unit 1, as explained below. Since with the broadening of the moving set of the power unit 1 and the torque transmitting second internal teeth 8 of the rotary piston 6 and the second outer teeth 9 of the shaft 2 is widened automatically, the high pressure level can be largely maintained and thus an increase in performance can be achieved. In the machines with propeller shaft drive between the rotor and the output shaft, this is not possible. Therefore, there is usually only allowed a lower pressure level at wider Laufsätzen with the stator 4 and the rotary piston 6. Motors with wide Laufsätzen run because of the higher absorption rate usually slower, so that the life of the bearings 10 and 11 is not a major problem.

Increasingly, a so-called "secondary regulation" is required on the market, and not only at hydraulic wheel drives, but increasingly also with hydraulically driven winches. The aim is to increase the speed range at the output without having to increase the capacity of the pump in relation to the flow rate. This is called rapid traverse operation, which usually occurs with reduced torque requirements. In Fig. 7 and in Fig. 8 is a hydraulic motor in longitudinal section or cross section according to the invention shown in which except the first power part 1 on a prolonged shaft end 44 of the shaft 2 a torsionally rigid coupled to the first power section 1 second, preferably narrower power section 46 is arranged with its own radial bearing 47, which can be operated separately via the ports 75 and 76 with working fluid, preferably from one and the same hydraulic pump. A proposal on the control of such a 2/3 stage motor with the first power section 1 and the second power section 46 is in Fig. 9 represented in the form of a hydraulic circuit diagram with exemplary performance data. By two separate 4/3-way valves 48 and 49 commercially available design can thus be driven at the same flow rate of a pump 83 up to three output speeds, as exemplified in Table 77. The forward and reverse positions of the 4/3-way valves are indicated by the letters F and R, respectively. It should be noted that the one motor stage that is switched to circulation and thus does not output torque, both on the displacement side, as well as especially on the intake side to be operated under high pressure, otherwise cavitation occurs at high speeds on the intake side. At the in Fig. 9 the scheme is taken into account. A throttle valve serves as a brake valve 87, in particular when driving downhill of the vehicle. By means of a valve 86, the operating state of the drive from operation D to neutral N can be switched.

In Fig. 10 is a further rotary engine according to the invention shown in longitudinal section, which of course also as a wheel motor according to Fig. 5 can be trained. In the embodiment, a hydraulically releasable spring-loaded working brake 50, designed as a disk brake, is arranged on a shaft extension 52. This work brake 50, the braking force is applied by means of springs 78, for example, in a hydrostatically driven winch for car or ship cranes the task to keep the full allowable rope load, which corresponds to the maximum pressure and thus the highest torque of the engine in the balance , without support hydraulic pressure on the engine. The load should be able to be sensitively manipulated up and down, so that the transition from the upward to the downward movement and vice versa, the pressure oil inflow on the rotary engine from primary to secondary must be switched. In this phase of change, the rotary engine has no torque because the pressure drops to zero. The spring-loaded work brake 50 takes over the holding torque at this moment and must therefore be designed so large that it can take over the maximum torque of the rotary piston engine. The size and number of springs 78 is to be sized accordingly, as well as the size and number of slats of the working brake 50th As can be seen from the Fig. 10 can recognize, a connectable via a separate connection 51 to the high-pressure pump high-pressure piston 79 is provided which is able to release the working brake 50, provided that the applied pressure on the high-pressure piston 79 by overcoming the spring forces of the springs 78 is large enough. In practice, it has been proven that this pressure must be between 8 and 12 bar, so that the load does not drop until the required support pressure is built up on the rotary piston engine.

It has been discussed a lot about whether such a large-sized brake for a high-torque motor, as it is present in the invention makes sense. The previous arrangement for such winch drives provides that instead of a rotary piston engine about a factor of 6 faster-running axial piston motor is used, which drives the sun gear of a planetary gear stage. Its torque is correspondingly smaller by a factor of 6. Between the axial piston motor and the planetary stage then the correspondingly also smaller by a factor of 6 dimensioned multi-disc brake of the same type is connected, similar to that in the Fig. 10 is shown. During operation of the winch, which must be driven in rapid traverse for reasons of saving time, this small brake runs relative to the housing, for example with 6 times the speed as the big brake according to the invention.

Wet-running multi-disc brakes have a particular advantage because they can be connected to the oil cooling system of the entire system through the oil passage. In addition, they are largely free of abrasion, so that the oil contamination is low. The disadvantage is that with oil-filled brake considerable, oil viscosity caused, loss-producing slip performance. According to Newton's law of shear stress in an oil gap, the slip power between two plates increases with the square of the relative velocity, and thus also between the running and stationary plates of a released brake. Assuming that when comparing the slip performance of a large brake according to Fig. 10 and a small brake described above, the oil viscosity, the thickness of the oil gap between the lamellae and the specific pressure on the lamellae by the spring forces are the same, then in the 6-fold faster running, small lamella brake, this slip performance is about 4 times as big as a slow running big brake according to Fig. 10 , It is thus obvious that apart from the more cost-effective solution, the compact version of a holding brake according to the invention, together with the high-torque motor described here, brings about an improvement in the overall efficiency of such a winch.

For the axial hydrostatic balance and the reduction of the axial running gaps to micrometer thickness between the control plate 22 and the rotary valve 3 on the one hand and between the rotary valve 3 and the axial balance piston 65 on the other hand, see Fig. 1 , very exact hydrostatic effective axial annular surfaces must be present. These are annular surfaces, which are defined by the respective average web diameter. They are in the FIGS. 1 . 5 . 6 . 7 and 10 not specially marked. However, as can be seen there, the diameter of the connecting holes 58a in the axial balance piston 65 and the subsequent holes in the rotary valve 3 are very small because the annular surface between the rotary valve 3 and the axial balance piston 65 is computationally relatively narrow. Although in the axial balance piston 65 very many such connection holes 58a are mounted on the circumference, so that the passage cross-section is relatively large. In contrast, in the rotary valve 3, the number of subsequent holes is very limited, because this must be based on the number of high-pressure windows 21a of the rotary valve 3.

This results in the problem that in these relatively small holes of the rotary valve 3, the flow velocity is very high. In hydraulics, the principle applies that at no point in the aggregate does the oil velocity in the high pressure range reach 10 to 12 m / s should exceed. Otherwise there will be strong turbulence, low static pressure according to Bernouilli's equation and possible cavitation damage to the canal walls. In addition, occurs at these locations at high flow rates on a disproportionate pressure drop, which reduces the performance and efficiency of the engine. Compared with known constructions, this disadvantage arises from the fact that in the inventive design, the rolling bearing right of the power unit has a large outer diameter. Thus, due to the system, the rotary valve 3 facing annular surface with the pressure windows 33a of the control plate 22 is relatively narrow (smaller diameter difference of the sealing webs). Accordingly, then the difference in the diameter of the mating ring surface between the rotary valve 3 and the axial balance piston 65 is smaller.

According to a development of the invention it is now proposed to move the counter-annular surface between the rotary valve 3 and the axial balancing piston 65 for the second annular space 58 in a smaller diameter range. In the event that the high pressure for the reverse direction of rotation is passed into the second annular space 58, the area of the annular surface must be the same for the balance of forces in this case as before. Thus, the difference in diameter of the sealing webs is considerably larger. In the Fig. 11 showing a longitudinal section through the advantageous development of the rotary piston engine according to the invention, these ratios are clearly shown. Starting from the central web diameter 95 and 96 of the control plate 22, see Fig. 11 and 12 , and the corresponding central web diameters 97 and 98 of the rotary valve 3, see Fig. 11 . 13 and 14 , which are represented by the dashed lines, initially remains the outer middle Stegdurchmesser 99 between the rotary valve 3 and the axial balance piston 65, see Fig. 11 and 13 , Same, because this together with the web diameter 97 causes the force balance on the rotary valve 3 in the event that the high pressure is supplied to the first annular space 56. In the other case, that the high pressure is supplied to the second annular space 58, the axial balance of the rotary valve 3, the new, in the inner diameter annular surface is responsible, which is determined by the new average web diameter 100 and 101.

The two left in Fig. 11 and 13 on the rotary valve 3 with their respective hydrostatic compensating forces acting annular surfaces are now completely separate from each other. This is done according to the invention by a screwed between the middle web diameters 99 and 100 circumferential axial relief groove 102, as in Fig. 11 and 13 can be seen. The axial relief groove 102 encircling the axial sliding surface 110 between the rotary valve 3 and the axial balance piston 65, see FIG Fig. 11 and 13 , thus lies between the first annulus 56 surrounding the rotary valve 3 and connected to the high pressure port 55 and the annular grooves 108 and 109 of the second annulus 58 connected to the low pressure port 57.

So that this relief groove 102 can really fulfill its separating function, it is connected to the leakage space 85 through the connection hole 103. The relief groove 102 and its communication hole 103 may be attached both in the rotary valve 3 and in the axial balance piston 65.

For a better understanding of the commutation of the rotary valve 3 are in the FIGS. 12 and 14 the required pressure window 33 a of the control plate 22 for supplying the tooth chambers of the power unit 1 and the high and Low-pressure window 21a and 21b shown in the rotary valve 3. The control mirror 104 of the control plate 22, Fig. 12 , Has between the pressure windows 33a evenly dimensioned dummy window 105, which are only a few tenths of a millimeter deep for a better isotropy of the lubricating film between the control mirror 104 and the rotary valve. 3

The advantages of this embodiment of the rotary piston engine according to the invention are considerable. A comparative study of the conditions according to FIGS. 1 . 5 . 6 . 7 , and 10 and the further embodiment according to the FIGS. 11 to 14 has shown that the diameter 106 of the bores in the rotary valve 3 can be increased approximately by a factor of 7/5. Since this is the narrowest point in this flow system, this improvement means that the oil flow and thus the speed of the rotary engine at the same oil speed at this point can be increased to about double. At the same time, the flow resistance is reduced and thus the pressure loss, so that the efficiency increases. Since at the same time the diameter 107 of the connecting bore 58a in the axial balancing piston 65 increases approximately in the same ratio, the flow loss is also reduced there and the number of required connecting bores 58a at the circumference of the axial balancing piston 65 can be smaller, which saves manufacturing costs. Also, the axial annular grooves 108 and 109 of the second annulus 58, see Fig. 11 and 13 , are enlarged in cross-section, which also contributes to a reduction in flow losses. Overall, this improvement means a considerable increase in performance and a higher overall efficiency of the rotary piston engine.

Of course it is possible; in the Fig. 11 to 14 illustrated development of the invention with features previously to combine described embodiments and equip, for example, a wheel motor or a winch drive with the last-described further development features.

Claims (27)

1. A hydrostatic, low-speed rotary cylinder engine, comprising

   • a power part (1) which acts as a drive and comprises

   ○ a central, stationary stator (4) having a first inner tooth system (5) with the number d of teeth,

   ○ a rotary piston (6) having a first outer tooth system (7) partly engaging the first inner tooth system (5) and having a number c of teeth and a second inner tooth system (8) having a number b of teeth and

   ○ a centrally mounted shaft (2) having a second outer tooth system (9) partly engaging the second inner tooth system (8) and having a number a of teeth,

   the rotary piston (6), for executing an orbital movement, being arranged eccentrically and dimensioned so that tooth chambers which can be supplied with working fluid and from which said fluid can be discharged form between the first inner tooth system (5) and the first outer tooth system (7),

   • an inlet and outlet part (70) for supplying working fluid to and discharging said fluid from the power part (1),

   • a disk-like rotary valve (3) for controlling the supply of the working fluid (2) and discharge of the working fluid from the tooth chambers,

   • an axial compensating piston (65) for sealing to prevent leakage at the rotary valve (3),

   • a toothed gear between an outer shaft tooth system (14) formed by a sun wheel (13) of the shaft (2) and a stationary inner toothed ring (92) as synchronous drives of the rotary valve (3) and

   • two roller bearings (10, 11) arranged directly adjacent on the shaft (2) on both sides of the power part (1),
   wherein

   • the rotary valve (3) is mounted so as to run concentrically with the shaft (2) and with the stator (4),

   • the toothed gear is arranged exclusively in the leakage oil region of the rotary cylinder engine and

   • the toothed gear is in the form of a planetary gear (80) having at least one planet carrier (91) which is non-rotatably connected to the rotary valve (3) and on which planet wheels (90) are arranged between the outer shaft tooth system (14) and the stationary inner toothed ring (92).
2. A hydrostatic, low-speed rotary cylinder engine, comprising

   • a power part (1) which acts as a drive and comprises

   ○ a central, stationary stator (4) having a first inner tooth system (5) with the number d of teeth,

   ○ a rotary piston (6) having a first outer tooth system (7) partly engaging the first inner tooth system (5) and having a number c of teeth and a second inner tooth system (8) having a number b of teeth and

   ○ a centrally mounted shaft (2) having a second outer tooth system (9) partly engaging the second inner tooth system (8) and having a number a of teeth,

   the rotary piston (6), for executing an orbital movement, being arranged eccentrically and dimensioned so that tooth chambers which can be supplied with working fluid and from which said fluid can be discharged form between the first inner tooth system (5) and the first outer tooth system (7),

   • an inlet and outlet part (70) for supplying working fluid to and discharging said fluid from the power part (1),

   • a disk-like rotary valve (3) for controlling the supply of the working fluid (2) and discharge of the working fluid from the tooth chambers,

   • an axial compensating piston (65) for sealing to prevent leakage at the rotary valve (3),

   • a toothed gear between an outer shaft tooth system (14) formed by a sun wheel (13) of the shaft (2) having a number w of teeth and a fourth inner tooth system (17) of a stationary inner toothed ring (28) having a number z of teeth as synchronous drive for the rotary valve (3), and

   • two roller bearings (10, 11) arranged directly adjacent on the shaft (2) on both sides of the power part (1),
   wherein

   • the rotary valve (3) is mounted so as to run concentrically with the shaft (2) and with the stator (4),

   • the toothed gear is arranged exclusively in the leakage region of the engine and

   • the toothed gear is in the form of an eccentric gear (30) having an eccentric (26) which is non-rotatably connected to the rotary valve (3), wherein the eccentric (26)

   • has a third inner tooth system (15) with a number x of teeth and a third outer tooth system (16) with a number y of teeth,

   • is arranged between the outer shaft tooth system (14) and the fourth inner tooth system (17) and

   • intermeshes with its third inner tooth system (15) with the outer shaft tooth system (14) of the shaft (2) and with its third outer tooth system (16) with the fourth inner tooth system (17) of the stationary inner toothed ring (28).
3. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 2, wherein

   • the eccentric gear (30) is in the form of a tumbling gear and

   • the eccentric is in the form of a disk-like eccentric (26) which is non-rotatably connected via a pot-like connecting part (27) to the rotary valve (3) via driver tooth systems (31, 32) in the speed ratio of 1:1.
4. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 4, wherein the numbers of teeth (a, b, c, d) of the power part (1) and the numbers of teeth (w, x, y, z) of the eccentric gear (30) fulfill the equation $$\frac{\frac{b}{a}\cdot d-c}{d-c}=\frac{\frac{x}{w}\cdot z-y}{z-y}$$

   

   
   and the result of this equation is a positive integer.
5. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 4, wherein the positive integer is equal to 3.
6. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 5, wherein the eccentric gear (30) is designed in such a way that the ratio of the revolutions per minute Ne of the eccentricity (20) of the eccentric gear (30) to the number of revolutions Nw of the shaft (2) according to the equation $$\frac{\mathit{Ne}}{\mathit{Nw}}=-\frac{w\cdot y}{x\cdot z-w\cdot y}$$

   

   
   is from -3 to -9.
7. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 4 to 6, wherein the number of teeth (a, b, c, d) of the power part (1) is a = 12, b = 14, c = 11 and d = 12 and the number of teeth (w, x, y, z) of the eccentric gear (30) is w = 12, x = 13, y = 23 and z = 24.
8. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 4 to 6, wherein the number of teeth (a, b, c, d) of the power part (1) is a = 12, b = 14, c = 11 and d = 12 and the numbers of teeth (w, x, y, z) of the eccentric gear (30) is w = 9, x = 10, y = 17 and z = 18.
9. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 2 to 8, wherein the common eccentricity (20) of the eccentric gear (30) is 0.013 to 0.015 times the mean reference circle diameter of control ports (21) in a control panel (22).
10. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 2 to 8, wherein the common eccentricity (20) of the eccentric gear (30) is 0.015 to 0.022 times the mean reference circle diameter of control ports (21) in a control panel (22).
11. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 2 to 10, wherein the number of teeth of the driver tooth systems (31, 32) between the eccentric (26) and the rotary valve (3) is twice as great as the number of teeth c of the first outer tooth system (7) of the rotary piston (6) of the power part (1).
12. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 1 to 11, wherein the first inner tooth system (5) of the stator (4) is formed by rotatably mounted rollers (81).
13. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 1 to 12, wherein a spring-loaded parking brake (42) which can be hydraulically released via a separate connection (43) is arranged on a shaft extension (41) of the shaft (2) on that side of the shaft (2) which is opposite the output side of the shaft (2).
14. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 1 to 12, wherein a spring-loaded working brake (50) which can be released via a separate connection (51) by the operating pressure of the rotary cylinder engine is arranged on a shaft extension (52) of the shaft (2) on that side of the shaft (2) which is opposite the output side of the shaft (2).
15. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 1 to 12, wherein a second power part (46) which is non-rotatably coupled to the first power part (1) and in particular has a separate radial bearing (47) for the lengthened shaft end (44) is arranged on a lengthened shaft end (44) of the shaft (2) on that side of the shaft (2) which is opposite the output side of the shaft (2).
16. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 15, wherein the specific intake of the second power part (46) is designed to be substantially smaller than that of the first power part (1).
17. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 15 or 16, wherein the first power part (1) and the second power part (46) are switchable by two separate 4/3-way valves (48) and (49).
18. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 17, wherein the power part (1, 46) switched in each case to revolution is switchable under feed pressure both on the divergent and on the convergent side of the intake or displacer system.
19. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 1 to 18, wherein a wheel flange (40) is arranged non-rotatably on the output side of the shaft (2) for directly driving a wheel which can be arranged on the wheel flange (40).
20. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 19, wherein the drive-side roller bearing (11) of the two roller bearings (10, 11) arranged directly adjacent on the shaft (2) on both sides of the power part (1) is arranged outside the leakage space (85) of the rotary cylinder engine with a permanent roller bearing grease fill, directly in the housing part (84) of the rotary cylinder engine.
21. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 19 or 20, wherein the wheel flange (40) is formed integrally with the shaft (2).
22. The hydrostatic, low-speed rotary cylinder engine as claimed in any of claims 1 to 12, wherein an all-round axial relief groove (102) is provided on an axial sliding surface (110) between the rotary valve (3) and the axial compensating piston (65), which relief valve is located between a first annular space (56) surrounding the rotary valve (3) and connected to a high-pressure connection (55) and annular grooves (108, 109) of a second annular space (58) connected to a low-pressure connection (57).
23. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 22, wherein the axial relief groove (102) is connected by a connecting bore (103) to the leakage space (85) of the rotary cylinder engine.
24. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 23, wherein the relief groove (102) and the connecting bore (103) thereof are arranged in the rotary valve (3).
25. The hydrostatic, low-speed rotary cylinder engine as claimed in claim 23, wherein the relief groove (102) and the connecting bore (103) thereof are arranged in the axial compensating piston (65).
26. A hydrostatic, low-speed wheel engine, comprising a hydrostatic rotary cylinder engine as claimed in any of claims 19 to 21, a wheel which can be driven directly by the hydrostatic rotary cylinder engine being arranged on the wheel flange (40).
27. A hydrostatic, low-speed winch drive, comprising a hydrostatic rotary cylinder engine as claimed in any of claims 19 to 21, a cable drum which can be driven directly by the hydrostatic rotary cylinder engine being arranged on the wheel flange (40).

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