EP3441613B1 - Hydrostatic gearwheel rotary piston machine - Google Patents

Description

The invention relates to a hydrostatic rotary piston gear machine based on the orbit principle according to the preamble of claim 1.

Such a hydrostatic rotary piston machine according to the preamble is in the EP 1 776 525 A1 described.

A hydrostatic rotary piston machine is known from the EP 0 367 046 A1 . In this known hydrostatic rotary piston machine, the working spaces between the toothing of the rotary piston and the stator are controlled by a drum-shaped rotary commutator. This rotary commutator is driven by a circular arc gear, which transfers the speed of the rotary piston to the speed of the rotary commutator 1: 1.

One advantage of this known hydrostatic rotary piston machine is that the rotary commutator and the circular arc gear are arranged between the roller bearings of the shaft and thus a relatively compact construction of the machine appears possible. However, due to the complexity and the existence of the drum-shaped rotary commutator and the circular arc gear between the bearings, the bearing spacing is relatively large, which in turn limits the performance potential of the machine.

This rotary commutator is mounted radially in the machine housing with the necessary radial running clearance. In the meantime, the working pressures of such rotary piston machines have increased to such an extent that such a constant, relatively large clearance no longer meets the volumetric requirements. In addition, the drive is via a Such a circular arc gear is afflicted with too much backlash, so that the rotational angle accuracy between the rotary piston and the rotary commutator suffers. An exact drive of the rotary commutator, also known as a rotary valve, and thus a high level of operational reliability at high working pressures over a long period of time are therefore not guaranteed without restrictions. Another disadvantage is the relatively complex structure of the machine.

From the EP 1 776 525 A1 , which is considered to be the closest prior art, a hydrostatic rotary piston machine is known in which the tapered roller bearings are arranged in the immediate vicinity of the rotary piston stator system to increase the performance potential. Here it is only possible with technically complex means to connect the rotary commutator, which is arranged outside the two bearings at a greater distance from the rotary piston-stator system, 1: 1 with the rotary piston in terms of speed kinematics. There, an eccentric gear and a planetary gear are proposed that can meet these kinematic conditions. This relatively complex construction requires the recreate specialist to have increased knowledge of the transmission analysis of such transmissions. In addition, the person skilled in the art cannot recognize the gear ratios by simply looking at them. These can only be proven mathematically. This increases the resistance among experts to the practical use of such constructions. Due to the necessary backlash in an eccentric wobble gear, the exact rotational position between the rotary piston and the rotary commutator is subject to compromises.

In a rotary piston machine, a centrally mounted shaft with external shaft teeth rotates centrally within a stator with a stationary in the housing Internal stator toothing, with a rotary piston arranged between the internal stator toothing and the external shaft toothing, the internal rotary piston toothing of which engages in the external shaft toothing with a different number of teeth and its external circular piston toothing in the internal stator toothing with a different number of teeth. The rotary piston executes a rotary movement around its own rotary piston axis, which runs parallel to, but not fixed to, the shaft and stator axes at a distance. When the shaft rotates, the rotary piston performs an orbital movement around the shaft axis in the area between the internal stator toothing and the external shaft toothing. The rotary piston axis thus performs a complex circular movement around the shaft and stator axis. Thus, the orbit movement, i.e. the rotational movement around an axis circling in a circular path, differs fundamentally from the simple rotational movements around the fixed axes of a gerotor machine. In a rotary piston machine, the tooth chambers filled with hydraulic fluid between the stator internal toothing and the rotary piston external tooth system also perform a circular movement around the shaft, while the tooth chambers in a gerotor machine are essentially fixed. In order to supply and dispose of the rotating tooth chambers with hydraulic fluid to generate the output, a rotary piston machine requires complex commutation by means of a rotary valve, also called a rotary commutator, which is driven at the speed of the rotary piston. Since such a rotary valve does not have to be used in a gerotor machine and the problem of its drive, as in particular in the EP 0 367 046 A1 and the EP 1 776 525 A1 described, does not exist, approaches from the field of gerotor machines for solving the problem with rotary piston machines are not or only partially applicable.

In the utility model DE 20 2014 006761 U1 describes a hydrostatic rotary piston machine based on the orbit principle, which is designed as a so-called ultra-slow runner. Compared to the previously known hydrostatic rotary piston machine according to the also described in the utility model EP 1 776 525 A1 , in which an internal gear is arranged between the rotary piston (rotor) and the shaft, whereby the shaft rotates faster than the rotary piston around its own axis, the ultra-slow runner described there differs mainly in that it has a 1: 1 rotary coupling is provided between the rotor and the shaft. Thus, the rotor rotates with the speed of the shaft, the displacement per shaft revolution is much larger and the output torque increases with the same working pressure. For 1: 1 speed and torque transmission between rotor and shaft, a wobble sleeve is arranged around the shaft, the external teeth of which mesh with the internal teeth of the rotary piston with the same number of teeth and internal teeth of the wobble sleeve with the same number of teeth engage in external teeth of the shaft. The disc-shaped rotary valve, which is used to dispose of and dispose of the tooth chambers, is also coupled to the shaft in a rotationally fixed manner at a speed ratio of 1: 1. In this way, the shaft, the rotary piston and the rotary valve always rotate at the same speed in the ultra-slow runner.

In U.S. 6,019,584 A a gerotor machine is described in which a centrically rotating shaft with a surrounding star-shaped rotor, which is arranged within a stator for performing an orbit movement, is connected 1: 1 in a rotationally fixed manner via a tumbling cylindrical coupling for torque transmission. A rotary valve is rotatably coupled 1: 1 to the shaft. This means that the shaft, rotor and rotary valve always rotate at the same speed here as well.

The object of the invention is to provide a hydrostatic gear rotary piston machine which is characterized on the one hand by a compact and relatively simple structure, on the other hand by a high performance potential, an exact drive of the rotary valve and high operational reliability even at high working pressures. The combination of these properties is currently considered a trade-off in the prior art.

This object is achieved by implementing the features of the independent claim. Features that develop the invention in an alternative or advantageous manner can be found in the dependent claims.

In order to use the advantages of these known constructions with regard to the high performance potential, according to the invention a disc-shaped rotary valve of the known design, in particular with axial play-compensated, is used between the two bearings and the drive of this rotary valve is improved by an almost play-free, direct, toothed cup-shaped wobble sleeve, which has a large diameter stable and continuous wave surrounds. In a further development, the internal gear between the shaft and the rotary piston is also optimized for the highest possible displacement per shaft revolution.

Due to these features according to the invention, the machine mentioned at the beginning can be used for much higher working pressures and the kinematic drive conditions for the disc-shaped rotary valve are much more precise, simpler and cheaper than the known machine mentioned. In addition, the invention reduces the dimensions and thus the weight of the machine by at least 15% can be reduced, which also contributes to a reduction in manufacturing costs. The compactness of the construction according to the invention is unsurpassed.

The invention represents a further development of the hydrostatic rotary piston machines known from the prior art, which is why the general basic structure of the hydrostatic rotary piston machine according to the invention, which is already essentially known from the prior art, is discussed below before the features according to the invention are explained.

The present invention is a rotary piston machine based on the orbit principle, with an internally toothed stator fixed in space, a rotary piston rotates around its own axis and this axis makes a circular movement in the opposite direction of rotation than the rotary piston. This means that in this overall movement the center distance line between the rotary piston and stator center executes a rotary movement in the opposite direction of rotation than the rotary piston around its own axis.

The hydrostatic rotary piston machine, which can also be referred to as a particularly slow-running high-torque rotary piston engine, comprises a power section acting as an output, which is arranged in the housing of the rotary piston machine or forms a logical section of the housing or the machine. The power section is essentially composed of a stationary, central stator, a movable rotary piston and a section of a centrally rotatably mounted shaft serving as an output, which engages with the rotary piston.

The stator has internal stator teeth with the number of teeth Z4. The rotary piston has a rotary piston external gear with a number of teeth Z3 and a rotary piston internal gear with a number of teeth Z2, which partially engages in the internal stator teeth. With its first external shaft toothing with a number of teeth Z1, the shaft partially meshes with the internal toothing of the rotary piston, the rotary piston being arranged and dimensioned eccentrically to the shaft axis in order to perform an orbit movement and the number of teeth Z1, Z2, Z3, Z4 in such a ratio are related to each other that tooth chambers that can be supplied and disposed of with working fluid are formed between the internal stator toothing of the stator and the external toothing of the rotary piston of the rotary piston.

The hydrostatic gear rotary piston machine according to the orbit principle thus comprises a central, stationary stator with internal stator teeth with the number of teeth Z4 and an eccentrically arranged rotary piston within the stator - for executing an orbit movement revolving around an eccentricity - one of which is partially inserted into the stator - has internal toothing engaging rotary piston external toothing with a number of teeth Z3, the difference in number of teeth between the number of teeth Z4 and the number of teeth Z3 being one. Z4 minus Z3 is therefore 1.

Inside the rotary piston there is a rotary piston internal toothing with a number of teeth Z2. A shaft rotatably mounted about a shaft axis is arranged centrically to the stator. On the shaft, in turn, an external shaft toothing with a number of teeth Z1, which partially engages the internal rotary piston toothing, is arranged. Tooth chambers are arranged radially between the stator internal teeth and the rotary piston external teeth and are of the stator internal toothing and the rotary piston external toothing limited radially in the radial direction and essentially sealed by the engagement of these toothings.

The supply and disposal of the tooth chambers with the working fluid is controlled via a disc-shaped rotary valve, which is mounted running centrically to the shaft and the stator. In other words, the rotary valve is mounted rotatably about the centric, geometric shaft axis extending along the center of the shaft. The rotary valve itself does not perform an orbit movement, but rotates around the shaft axis.

By means of the rotary valve, the supply and disposal of the tooth chambers with the working fluid can be controlled commutatingly by rotating the rotary valve via pressure windows arranged essentially radially in the rotary valve in such a way that the working fluid from the pressure connection into a first part of the tooth chambers with the working pressure and out of a second Part of the tooth chambers is passed out to a low-pressure connection, so that the working pressure in the first part of the tooth chambers leads to an orbit movement of the rotary piston and displaces the working fluid from the second part of the tooth chambers, causing the shaft to rotate, and vice versa.

In other words, the disc-shaped rotary valve can be rotated centrically about the shaft axis and is mounted running centrically to the shaft and to the stator. The disk-shaped rotary valve is for the commutating control of the supply and disposal of the tooth chambers with the working fluid for introducing working fluid into a first part of the tooth chambers with the working pressure and for discharging the working fluid from a The second part of the tooth chambers is designed to generate the output and is connected to the tooth chambers.

The shaft is supported by radial bearings, especially roller bearings. A main bearing is arranged on a shaft output side of the shaft. A secondary bearing is arranged on an opposite side of the shaft, which is at the other end of the shaft - opposite the shaft output side. The main bearing and / or the secondary bearing are designed in particular as radial bearings, as roller bearings, as roller bearings or preferably as tapered roller bearings, in particular in an X arrangement.

Due to the aforementioned structure of the hydrostatic rotary piston machine, a continuous shaft with large shaft diameters and high torsional strength can be used. It is thus possible to expose both shaft ends to a high torque flow and, for example, to use both shaft ends as an output, or one shaft end as an output and the other shaft end to connect a brake or a second drive, whereby the entire drive unit can be made compact.

In order to axially seal the disc-shaped rotary valve against leakage of the working fluid and to press it with its pressure windows with sufficient pressure, in particular against the control plate with its windows, which lead to the tooth chambers via control plate channels, a non-rotatable hydrostatic axial compensating piston is preferably used for axial backlash freedom of the disc-shaped rotary valve is provided, which acts axially on the rotary valve. The axial compensating piston is arranged around the shaft.

The hydrostatic rotary piston gear machine has a housing that is logically divided into several parts, i.e. with regard to individual functional sections. These parts are not necessarily to be understood as separate or separable pieces, but rather as logical sections. It is thus possible to form several parts in one piece or individual parts in several parts. In a preferred embodiment, however, the individual housing parts are each designed as a separable housing piece.

The housing is divided into at least an output-side housing part, an inlet and outlet housing part and a power part located axially between them in the axial direction, i.e. along the shaft axis, and in particular a control plate arranged axially between the inlet and outlet housing part and the power part.

The housing part on the output side includes the main bearing. The main bearing is arranged radially between the housing part on the output side and the shaft output side of the shaft. The shaft output side of the shaft and the shaft end there are led out of the housing from the housing part on the output side.

Viewed in the axial direction, the inlet and outlet housing parts are located on the side of the housing opposite the housing part on the output side. This is used for supplying and disposing of the power unit with the working fluid, the working fluid being able to be fed to the rotary piston machine at a working pressure via a pressure connection that serves as an inlet. The inlet and outlet housing part includes the disc-shaped rotary valve, which is rotatably mounted therein about the shaft axis. In addition, the inlet and outlet housing part includes connections and Channels which are designed for the supply and disposal of the disc-shaped rotary valve with the working fluid. In addition, the inlet and outlet housing parts include, in particular, the compensating piston.

The power part, seen in the axial direction between the output-side housing part and the inlet and outlet housing part, acts as a drive for the machine and includes the stator, the rotary piston, the tooth chambers and the external shaft toothing.

The control plate comprises windows and control plate channels in order to guide the working fluid essentially in the axial direction from the disc-shaped rotary valve to the tooth chambers.

The housing parts are preferably braced with one another by means of axially extending screws.

According to the invention, the secondary bearing in the inlet and outlet housing part and the disc-shaped rotary valve are arranged between the power part and the secondary bearing, viewed in the axial direction. The power section and the rotary valve, and in particular the hydrostatic axial compensating piston and a cup-shaped sleeve, are located axially between the main bearing and the secondary bearing.

This cup-shaped sleeve, which is also referred to below as a wobble sleeve, is arranged radially around the shaft and extends in the axial direction at a radial distance therefrom. The cup-shaped sleeve surrounds the shaft axis and the shaft tumbling, since it is rotatably mounted on one end about the shaft axis, and rotatably on the other end about the rotary piston axis. The sleeve extends, viewed in the axial direction, between the power unit and the input and output Outlet housing part in the housing and is in particular passed freely through the control plate.

The sleeve has a first sleeve external toothing and a second sleeve external toothing. The first sleeve external toothing engages with a number of teeth Z5 on the side of the power section in the rotary piston internal toothing and meshes with it. The number of teeth Z5 of the first sleeve external toothing is preferably equal to the number of teeth Z2 of the rotary piston internal toothing. The second sleeve external toothing engages with a number of teeth Z6 on the side of the inlet and outlet housing part in a rotary valve internal toothing of the rotary valve and meshes with these, the rotary valve internal toothing having the number of teeth Z7. The number of teeth Z6 of the second sleeve external toothing is preferably equal to the number of teeth Z7 of the rotary valve internal toothing.

The sleeve and its teeth are designed for a tumbling 1: 1 rotary coupling between the rotary piston and the disc-shaped rotary valve and enable the rotary valve to be operated at the speed of the rotary piston, which, however, is in relation to the rotary valve around another axis rotating in space - the rotary piston axis - turns to drive.

The difference in number of teeth between the number of teeth Z2 and the number of teeth Z1 is preferably equal to 2. Thus, Z2 minus Z1 equals 2.

The cup-shaped sleeve can be cylindrical or conical, but also have any other shape which at least partially surrounds the shaft and has the said first sleeve external toothing at one end and the said second sleeve external toothing at the other end. For example, the cup-shaped sleeve can also have a lattice structure.

In a preferred embodiment, which is explained in more detail below, the number of teeth Z3 has the values 10:11 or 11:12 or 12:13 for the number of teeth Z4, and / or the number of teeth Z1 has the values 15 for the number of teeth Z2 : 17 or 16:18 or 17:19. The number of teeth Z3 for the number of teeth Z4 preferably has the value 11:12 and the number of teeth Z1 for the number of teeth Z2 has the value 16:18.

In a further development, the hydrostatic axial compensating piston is arranged between the disc-shaped rotary valve and the auxiliary bearing in the inlet and outlet housing parts, seen in the axial direction.

At high working pressures, the volumetric efficiency is always an important variable, as it has a decisive influence on the overall efficiency. The factors influencing the volumetric efficiency are, among other things, the axial leakage gap between the rotary piston and the side walls. For example, if the design of the axial running clearance is too small, the machine runs the risk of the rotary piston jamming axially at a higher temperature, with the result that there is a risk of seizure and, at the same time, the mechanical efficiency drops. A resilience of at least one side wall in the machine housing is therefore desirable. Without special measures, however, this resilience would result in the internal hydraulic axial forces increasing the axial play again because of the lower rigidity of this side wall.

According to the invention it is therefore proposed that seen in the axial direction between the power part and the inlet and outlet housing part, adjacent and adjacent to Power part, a control plate - with windows and control plate channels between the tooth chambers and the disc-shaped rotary valve - is arranged for supplying and disposing of the tooth chambers with the working fluid, and this control plate is designed to be flexible. The latter is achieved in particular in that the thickness in the axial direction of the control plate and / or its material are such that the axial running clearance of the rotary piston between the output-side housing part and the inlet and outlet housing part decreases or remains the same as the working pressure of the working fluid increases.

The invention therefore provides for this control plate to be made much thinner because of the required axial flexibility, but to support it on the opposite side with sufficiently large hydraulic compensation surfaces. These are available when the axially acting compensation surfaces on the rotary valve are large enough. Correspondingly, the compensation surface of the compensating piston should then also have an overcompensated size. If these hydraulic compensation surfaces on the control plate are sufficiently overcompensated, e.g. by 10 to 15%, there is a possibility that the axial play of the rotary piston between its side walls will decrease as the working pressure increases. Since the axial leakage current decreases according to the Hagen-Poiseuille law with the third power of the leakage gap reduction, the volumetric efficiency can be improved, especially at high working pressure.

The flexibility of the control plate can also be influenced by the selection of the material used for it. In such cases, plain bearing materials such as lead bronze or brass alloys are popular. There are also high-strength ones Aluminum plain bearing alloys are an option. At the same time, these have the great advantage that emergency running properties significantly improve the sliding conditions. The significantly lower modulus of elasticity of these materials increases their above-mentioned compliance.

The flexibility of the control plate can be particularly effectively influenced by the selection of the thickness of the plate. According to the theory of plates and shells, plate stiffness is proportional to the cube of the plate thickness, inversely proportional to the modulus of elasticity of the plate material. Thus, the deflection of the circular ring plate clamped at the edge can be calculated relatively easily, provided the differential force is known. What is important here is the deflection on the inner edge of the circular ring plate that is involved here.

Due to the large distance between the tapered roller bearings, this concept is predestined for a design as a hydrostatic wheel motor, since the output-side tapered roller main bearing with its high load rating must still absorb the wheel load in addition to the tooth force on the external shaft toothing from the rotary piston. The shaft output side of the shaft is therefore designed in an advantageous embodiment as a cone for fastening a wheel flange by means of an axial nut to form a compact wheel motor.

In an advantageous embodiment, the internal stator toothing is formed by rotatable, in particular circular-cylindrical rollers, which leads to further increased pressure output and excellent service life and as a result of which the start-up efficiency and also the mechanical-hydraulic efficiency can be increased considerably.

Furthermore, in a further development, the shaft external toothing of the shaft is designed conically with a smaller diameter amount, viewed in the axial direction, on its shaft output side, as explained in more detail in the exemplary embodiments.

In addition, one aspect of the invention provides that a tooth tip shortening is provided at the tooth tip of the external shaft toothing, in that the tooth flank radius is calculated to be larger at the points and on the tooth tip where the teeth come out of engagement than at the point of the deepest tooth engagement, which is also the case will be discussed in more detail in the exemplary embodiments.

Because of the particularly compact design of the rotary piston machine according to the invention, particularly its axial dimension, there is the possibility of coupling two machines axially in a torque-effective manner without the unit becoming excessively long. The advantage of such an arrangement is that a two- to three-stage hydraulic motor is created by means of a relatively simple slide valve control. When used as a drive for a hydrostatic cable winch in hoists, for example, the speed of the empty hook can be increased. When used as a drive for a hydrostatic mobile vehicle, this allows the vehicle speed to be increased when driving empty. In both cases, a time saving can be achieved, so that the hydraulic oil quantity of the delivery pump can be kept constant. In this case, the output speed of the rotary piston machine can be spread by a factor of up to 3 without great structural effort.

Therefore, a development of the invention comprises a hydrostatic rotary piston gear machine in the form of a Gear double rotary piston machine, which is composed of two hydrostatic gear rotary piston machines coupled to one another. These two hydrostatic gear rotary piston machines each correspond to one of the hydrostatic gear rotary piston machines according to the invention. The shaft opposite side of the shaft of the first hydrostatic gear rotary piston machine is axially coupled to the shaft opposite side of the shaft of the second hydrostatic gear rotary piston machine, in particular by means of a sleeve, and is connected with effective torque. The first hydrostatic gear rotary piston machine and the second hydrostatic gear rotary piston machine have, in particular, different absorption quantities, in particular differently dimensioned tooth chambers. In particular, a first splined shaft / toothed hub toothing on the opposite side of the shaft of the shaft of the first hydrostatic geared rotary piston machine and a corresponding second splined shaft / toothed hub toothing of the second hydrostatic toothed / splined piston machine are each connected to the sleeve in a torque-effective manner. This sleeve is in particular supported radially by a common roller bearing in the inlet and outlet housing part of the first hydrostatic gear rotary piston machine and in the inlet and outlet housing part of the second hydrostatic gear rotary piston machine. Here, the roller bearing is designed as a centering of the first hydrostatic gear rotary piston machine with the second hydrostatic gear rotary piston machine.

The hydrostatic rotary piston gear machine according to the invention is described in more detail below, purely by way of example, with the aid of specific exemplary embodiments shown schematically in the drawings, further advantages of the invention also being discussed.

Figur 1

einen Längsschnitt durch eine erste Ausführungsform der hydrostatischen Zahnrad-Kreiskolbenmaschine;

Figur 2

einen Querschnitt durch den Leistungsteil der hydrostatischen Zahnrad-Kreiskolbenmaschine nach Figur 1;

Figur 3

eine Tabelle mit unterschiedlichen Zähnezahlverhältnissen;

Figur 4

einen Längsschnitt durch eine als Radmotor ausgebildete zweite Ausführungsform der Erfindung; und

Figur 5

einen Längsschnitt durch eine als Zahnrad-Doppelkreiskolben-Maschine ausgebildete dritte Ausführungsform der Erfindung.

Show in detail:

Figure 1

a longitudinal section through a first embodiment of the hydrostatic gear rotary piston machine;

Figure 2

a cross section through the power part of the hydrostatic gear rotary piston machine Figure 1 ;

Figure 3

a table with different tooth number ratios;

Figure 4

a longitudinal section through a second embodiment of the invention designed as a wheel motor; and

Figure 5

a longitudinal section through a designed as a gear double circular piston machine third embodiment of the invention.

Figure 1 shows a longitudinal section through a first embodiment of the hydrostatic gear rotary piston machine, while the Figure 2 shows a cross section through the power part of this machine. The two Figures 1 and 2 are described together below.

The hydrostatic gear rotary piston machine according to the orbit principle has a central, stationary stator 2 provided with internal stator teeth 1 with the number of teeth Z4 equal to 12, as well as an eccentrically arranged rotary piston 4 within the stator 2 for executing an orbit movement that revolves around an eccentricity e , the one partially engaging in the stator internal toothing 1 Rotary piston external toothing 3 with a number of teeth Z3 equal to 11. The difference in number of teeth between the number of teeth Z4 equal to 12 and the number of teeth Z3 equal to 11 is 1, as in FIG Figure 2 evident. The internal stator toothing 1 is designed as rotatable rollers.

In the interior of the rotary piston 4, a rotary piston internal toothing 5 is formed with a number of teeth Z2. A shaft 6 is arranged centrally with respect to the stator 2 and is mounted rotatably about a geometric shaft axis 52 in a tapered roller main bearing 9 and a tapered roller secondary bearing 11. Shaft external toothing 7 with a number of teeth Z1 is formed on shaft 6 and partially engages in rotary piston internal toothing 5. The geometrical shaft axis 52 is thus also the geometrical axis of the external shaft toothing 7, the stator 2 and the internal stator toothing 1. The difference in number of teeth between the number of teeth Z2 and the number of teeth Z1 is 2.

The tooth chambers 53a, 53b formed by the difference in number of teeth between the number of teeth Z4 and the number of teeth Z3 between the stator internal gear 1 and the rotary piston external gear 3 are radially separated from the stator internal gear 1 and the rotary piston external gear 3, as in Figure 2 shown, and axially from a wall 58 on the output side and opposite from a control plate 25, as in FIG Figure 1 shown, limited essentially sealed.

The stator 2, the rotary piston 4, the tooth chambers 53a, 53b and the external shaft toothing 7 form the power part 51 of the housing 24, which acts as a drive.

On a shaft output side 8 of the shaft 6, the tapered roller main bearing 9 is arranged, while on a the shaft output side 8 at the other end of the shaft 6 opposite shaft opposite side 55, the tapered roller sub-bearing 11 is located.

In Figure 1 the extremely stable design of the shaft 6 is shown, the output-side tapered roller main bearing 9 being selected with a particularly high load rating. This is arranged very close to the tooth engagement between the rotary piston 4 and the external shaft toothing 7 of the shaft 6 for the smallest possible deflection of the shaft 6 due to the tooth force, since the tapered roller sub-bearing 11 is placed at the greatest possible distance therefrom. At maximum working pressure, the amount of shaft deflection at the point of tooth meshing should not exceed 15 to 20 micrometers. So that at maximum pressure of the working fluid no edge carrier arises between the rotary piston internal toothing 5 of the rotary piston 4 and the shaft external toothing 7 of the shaft 6 as a result of the deflection of the shaft 6, according to the invention there is the possibility that the shaft external toothing 7 is ground slightly conically with something smaller diameter amount, seen in the axial direction, on its output side in the direction of the shaft output side 8.

Right in Figure 1 shown is an inlet and outlet housing part 54 of the housing 24, which is an output-side housing part 10, on the left in Figure 1 , viewed in the axial direction opposite. The inlet and outlet housing part 54 contains a disk-shaped rotary valve 12 as well as connections 21 and 22 and channels 56 for the supply and disposal of the disk-shaped rotary valve 12 with working fluid.

The disc-shaped rotary valve 12 can be rotated centrically about the shaft axis 52 and is mounted running centrically to the shaft 6 and to the stator 2 in the inlet and outlet housing part 54 between the power unit 51 and the tapered roller sub-bearing 11 and is used for the commutating control of the supply and disposal of the tooth chambers 53a, 53b with the working fluid for guiding working fluid into a first part 53a of the tooth chambers with a working pressure and for guiding the working fluid out of a second part 53b of the tooth chambers to generate the output.

In other words, the rotary valve 12 is used to connect a first part 53a of the circling tooth chambers to one of the two connections 21 and 22 for supplying these tooth chambers 53a with the working fluid under working pressure and a second part 53b of the circling tooth chambers to the other of the two connections 21 and 22 designed to discharge the working fluid from these tooth chambers 53b and connected to the tooth chambers 53a, 53b via the channels 56, respectively.

Also arranged in the inlet and outlet housing part 54 is a hydrostatic axial compensating piston 17 for the disc-shaped rotary valve 12 to have no axial play.

Viewed in the axial direction between the power part 51 and the inlet and outlet housing part 54, the control plate 25 is arranged. This control plate 25 has windows 26 and control plate channels 57 between the tooth chambers 53a, 53b and the pressure windows 27 formed in the disc-shaped rotary valve 12 for supplying and removing the tooth chambers 53a, 53b with the working fluid.

The thickness d in the axial direction of the control plate 25 and its material are such that the axial running clearance of the rotary piston 4 between the output-side housing part 10 and the inlet and outlet housing part 54 decreases or remains the same as the working pressure of the working fluid increases.

Like in Figure 1 As shown, the tapered roller main bearing 9 is arranged in the output-side housing part 10, from which the shaft output side 8 of the shaft 6 is led out of the housing 24, while the tapered roller sub-bearing 11 is located in the inlet and outlet housing part 54.

Seen in the axial direction between the output-side housing part 10 and the control plate 25, the power part 51 acting as a drive is arranged with the stator 2, the rotary piston 4, the tooth chambers 53a, 53b and the external shaft toothing 7.

In order to be able to drive the rotary valve 12, which is rotatable about the shaft axis 52, at the speed of the rotary piston 4 which is rotatable about the rotary piston axis 50 - despite the offset of the axes of rotation by the eccentricity e - a cup-shaped sleeve 13 is arranged between the rotary piston 4 and the rotary valve 12. The cup-shaped sleeve 13 surrounds the shaft 6 at a radial distance and wobbles about the shaft axis 52 when the rotary piston 4 rotates. As in FIG Figure 1 As shown, it extends, viewed in the axial direction, between the power part 51 and the inlet and outlet housing part 54 in the housing 24 and is passed through a corresponding recess in the control plate 25. The wobble sleeve 13 has a first sleeve external toothing 14 with a number of teeth Z5 on the side of the power section 51, i.e. in the direction of the output side, and a second sleeve external toothing 15 with a number of teeth Z6 on the side of the inlet and outlet housing part 54, i.e. in Direction opposite to the output side. The first sleeve external toothing 14 meshes with the rotary piston internal toothing 5, the number of teeth Z2 of which is equal to the number of teeth Z5. The second sleeve external toothing 15 meshes with a rotary valve internal toothing 16 with a number of teeth Z7 of the rotary valve 12 corresponding to the number of teeth Z6.

In the exemplary embodiment shown, the cup-like wobble sleeve 13 performs a wobble movement with a wobble angle of approximately 5.5 degrees during the orbital movement of the rotary piston 4. The backlash between the teeth of the first sleeve external toothing 14 and the rotary piston internal toothing 5 and between the teeth of the second sleeve external toothing 15 and the rotary valve internal toothing 16 should preferably be as small as possible for an exact commutating control of the power unit from the rotary piston External teeth 3 and the stator internal teeth 1. Therefore, the teeth of the first sleeve external teeth 14, the second sleeve external teeth 15 and the rotary valve internal teeth 16 are very narrow.

So that the frictional loss torque of the rotary valve 12 is as small as possible, blind depressions should also be provided between the windows 26 of the control plate 25, seen in the circumferential direction. The arrangements of these windows 26 of the control plate 25 and the pressure windows 27 of the rotary valve 12 are shown in simplified form in the drawing and are generally known to those skilled in the art.

The stator internal teeth 1 and the rotary piston external teeth 3 of the gerotor set and the rotary piston internal teeth 5 and shaft external teeth 7 of the Internal gear between the shaft 6 and the rotary piston 4 are in Figure 2 shown. Since one and the same eccentricity e, i.e. the same axial distance between the shaft axis 52 and the rotary piston axis 50, must apply to both internal gears, the difference in the number of teeth in the shaft eccentric internal gear is two. So that the displacement of the machine and the shaft diameter are as large as possible, the number of teeth is also chosen to be large.

The tooth pitch and, accordingly, the module of the teeth are based on the amount of eccentricity e of the entire running set, which at the same time means the center distance for both internal gears. The design of the internal gear between the shaft 6 and the rotary piston 4 has a significant influence on the absorption volume per shaft revolution and thus on the hydraulic performance of the machine.

Figure 3 shows a table that makes these relationships clear. The essential size for the absorption volume is the so-called conveying surface Ao of the rotor toothing between the rotary piston external toothing 3 and the stator internal toothing 1. The calculation is based on the so-called stationary gear, in which the center distance line or the eccentricity e is stationary in space. In this state of motion, the rotors do not run as a rotary piston gear, but rather as a rotary piston gear. Both wheels rotate around their own centers and form a normal internal gear machine.

dabei bedeuten

• Dk1 der Kopfkreisdurchmesser des Innenläufers,
• Dk2 der Kopfkreisdurchmesser des Außenläufers,
• e der Achsabstand oder die Exzentrizität des Getriebes.

The basic delivery area Ao (delivery rate per cm width of the rotor) per revolution of the inner rotor (later rotary piston) is then calculated with sufficient accuracy using the formula: $$\mathrm{Ao}=\left(\mathrm{pi}/\mathrm{<figure-callout\; id="1"\; label="4th\; \cdot \; Dk"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">4th</figure-callout>}\right)\cdot \left[\mathrm{Dk}\right]\left[1^2-{\left(\mathrm{Dk}2-2\mathrm{<figure-callout\; id="2"\; label="e"\; filenames="imgf0002.png"\; state="\{\{state\}\}">e</figure-callout>}\right)}^2\right]\left[{\mathrm{<figure-callout\; id="2"\; label="cm"\; filenames="imgf0002.png"\; state="\{\{state\}\}">cm</figure-callout>}}^2\right]$$

mean thereby

• Dk1 is the tip diameter of the inner rotor,
• Dk2 is the tip diameter of the external rotor,
• e the center distance or the eccentricity of the gearbox.

As you can see, Ao is independent of the number of teeth. For a comparative investigation of the displacement, it is assumed that the outside diameter of the machine is the same in every case, thus also the basic conveying area Ao.

$$\mathrm{Nw}/\mathrm{Nk}=\left[\mathrm{Z}4\cdot \left(\mathrm{Z}2/\mathrm{Z}1\right)-\mathrm{Z}3\right]/\left[\mathrm{Z}4-\mathrm{Z}3\right]$$

Dabei bedeuten

• Z1 die Zähnezahl der Wellen-Außenverzahnung 7,
• Z2 die Zähnezahl der Kreiskolben-Innenverzahnung 5,
• Z3 die Zähnezahl der Kreiskolben-Außenverzahnung 3,
• Z4 die Zähnezahl der Stator-Innenverzahnung 1,
• Ne die Drehzahl der Exzentrizität e (negativ),
• Nk die Drehzahl des Kreiskolbens 4 (positiv),
• Nw die Drehzahl der Welle 6 (positiv).

In the case of the rotary piston machine, the external rotor (now referred to as the stator 2) is stationary in space, the internal rotor becomes the rotary piston 4 and describes an "orbit" movement around the shaft center point with the distance e. The rotary piston 4 continues to rotate in the positive direction, as does the shaft 6. However, the eccentricity e rotates in the opposite direction of rotation, that is to say at a negative speed. Then the speed ratios are according to the laws of epicyclic gears: $$\mathrm{No}/\mathrm{Nk}=\left(-\mathrm{<figure-callout\; id="3"\; label="Z"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">Z</figure-callout>}3/1\right);$$

$$\mathrm{Nw}/\mathrm{Nk}=\left[\mathrm{Z}\mathrm{4th}\cdot \left(\mathrm{<figure-callout\; id="2"\; label="Z"\; filenames="imgf0002.png"\; state="\{\{state\}\}">Z</figure-callout>}2/\mathrm{Z}1\right)-\mathrm{<figure-callout\; id="3"\; label="Z"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">Z</figure-callout>}3\right]/\left[\mathrm{Z}\mathrm{4th}-\mathrm{<figure-callout\; id="3"\; label="Z"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">Z</figure-callout>}3\right]$$

Mean

• Z1 is the number of teeth on the external shaft toothing 7,
• Z2 the number of teeth of the internal gearing of the rotary piston 5,
• Z3 the number of teeth of the rotary piston external toothing 3,
• Z4 the number of teeth of the stator internal toothing 1,
• Ne is the speed of the eccentricity e (negative),
• Nk is the speed of the rotary piston 4 (positive),
• Nw is the speed of shaft 6 (positive).

• bei Z4-Z3 = 1
• Ak = Z4·Ao

With this rotary piston movement, the delivery area Ak per rotary piston revolution increases dramatically according to the formula: $$\mathrm{Ak}=\left[-\left(\mathrm{No}/\mathrm{Nk}\right)+1\right]\cdot \mathrm{Ao}=\left(\mathrm{<figure-callout\; id="3"\; label="Z"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">Z</figure-callout>}3+1\right)\cdot \mathrm{Ao}$$

• with Z4-Z3 = 1
• Ak = Z4 * Ao

Since the shaft 6 rotates faster than the rotary piston 4, the conveying surface Aw of the machine per revolution of the shaft 6 is: $$\mathrm{Aw}=\left[\mathrm{Ak}/\left(\mathrm{Nw}/\mathrm{Nk}\right)\right]\cdot \mathrm{Ao}=\left\{\left[\mathrm{<figure-callout\; id="3"\; label="Z"\; filenames="imgf0001.png,imgf0002.png"\; state="\{\{state\}\}">Z</figure-callout>}3+1\right]/\left[\mathrm{Nw}/\mathrm{Nk}\right]\right\}\cdot \mathrm{Ao}$$

In the table according to the Figure 3 three different designs of the gerotor set Z3 / Z4 and three different designs of the rotary piston internal gearbox Z1 / Z2 are assumed, so that 9 different results are obtained for the shaft conveying area Aw per shaft revolution, which are given in the last column of the table. As you can see, the values for Aw increase with an increasing number of teeth on the external shaft toothing 7.

Thus the total displacement of the machines per shaft revolution is: $$\mathrm{qw}=\mathrm{Aw}\cdot \mathrm{B.},\mathrm{if}\mathrm{B.}\mathrm{the}\mathrm{broad}\mathrm{of}\mathrm{Running\; set}\mathrm{is}\mathrm{.}$$

The in the table in Figure 3 The framed design corresponds to the representation in the figures.

The small difference in number of teeth of only two teeth between the rotary piston internal toothing 5 and the shaft external toothing 7 can be at the points 28 and 29 of the tooth tip of the shaft external toothing 7, where the Toothings disengage, tooth tip engagement disorders occur ( Figure 2 ). In order to avoid this, it is therefore expedient that a tooth tip shortening is provided at these points of the tooth tip of the shaft external toothing 7. This can be done in a simple manner in that in the calculation program for the tooth shape of the shaft external toothing 7 at these points the tooth flank radius 30 of the rotary piston internal toothing 5 is calculated to be somewhat larger than at the point of the deepest tooth engagement 31 Transition of this tooth flank radius 30 from the point of deepest tooth engagement 31 to points 28 and 29 at which the rotary piston internal toothing 5 disengages.

Figure 4 shows a longitudinal section through a second embodiment of the invention designed as a wheel motor as a particularly compact construction.

This second embodiment differs in particular from the first embodiment in that the shaft output side 8 of the shaft 6 is designed as a cone 37 and a wheel flange 18 is attached to this cone 37 by means of an axial nut 19 to form a compact wheel motor.

Figure 5 shows a longitudinal section through a third embodiment of the invention designed as a gear double circular piston machine. This two to three-stage hydrostatic rotary piston machine is a combination of a first hydrostatic gear rotary piston machine 40 with a large displacement per revolution and a second hydrostatic gear rotary piston machine 41 with a smaller displacement per revolution. The first and second rotary hydrostatic gear machines 40 and 41, respectively correspond in terms of their essential features to the hydrostatic rotary piston machine according to the invention described at the beginning, but they have tooth chambers 53a, 53b of different sizes with different axial extensions, as in FIG Figure 5 shown.

The respective opposite shaft sides 55 of the shafts 6 of both machines 40 and 41 are connected to one another in a torque-effective manner via a sleeve 42 via toothed shaft and toothed hub teeth 43a and 43b, the sleeve 42 being connected radially by a common roller bearing 44 in the inlet and outlet housing part 54 of the first hydrostatic gear rotary piston machine 40 and is mounted in the inlet and outlet housing part 54 of the second hydrostatic gear rotary piston machine 41. The roller bearing 44 is used in an advantageous manner to center the two machines 40 and 41. By means of the screws 46, both units are screwed in the axial direction.

As can be seen from the drawing, the length of this multi-stage hydrostatic rotary piston machine is only about twice the diameter.

Claims (14)

1. Hydrostatic gearwheel rotary piston machine according to the orbit principle, having

   • a centric, fixed stator (2) provided with a stator internal toothing (1) and with the number of teeth Z4 and,

   • a rotary piston (4) arranged eccentrically inside the stator (2) for carrying out an orbital movement circling about an eccentricity (e), which rotary piston (4) comprises a rotary piston external toothing (3) with a number of teeth Z3 partially engaging in the stator internal toothing (1),

   • a difference in the number of teeth 1 between the number of teeth Z4 and the number of teeth Z3,

   • a rotary piston internal toothing (5) arranged inside the rotary piston (4) with a number of teeth Z2,

   • a shaft (6) arranged centrically to the stator (2) and mounted rotatably about a shaft axis (52),

   • a shaft external toothing (7) with a number of teeth Z1 arranged on the shaft (6) and partially engaging in the rotary piston internal toothing (5),

   • tooth chambers (53a, 53b) which are arranged radially between the stator internal toothing (1) and the rotary piston external toothing (3) and which are radially delimited by the latter,

   • a main bearing (9) arranged on a shaft output side (8) of the shaft (6),

   • a secondary bearing (11) arranged on a shaft opposite side (55) opposite to the shaft output side (8) at the other end of the shaft (6),

   • a disc-shaped rotary valve (12) which is centrically rotatable about the shaft axis (52) and is centrically running with respect to the shaft (6) and the stator (2), and which is adapted to commutatively control the supply and discharge of the working fluid to the tooth chambers (53a, 53b) with the working fluid for introducing working fluid into a first part (53a) of the tooth chambers with a working pressure and for discharging the working fluid out of a second part (53b) of the tooth chambers for generating the output and is connected to the tooth chambers (53a, 53b),

   • a hydrostatic axial balance piston (17) for an axial freedom from play of the disc-shaped rotary valve (12), and

   • a housing (24) having

   - a housing section (10) on the output side, which comprises the main bearing (9) and from which the shaft output side (8) of the shaft (6) is guided out of the housing (24),

   - an inlet and outlet housing section (54), which is opposite the housing section (10) on the output side, as viewed in the axial direction, and which comprises the disc-shaped rotary valve (12) having connections (21, 22) and channels (56) formed for supplying and discharging the working fluid to the disc-shaped rotary valve (12), and in particular the balance piston (17), and

   - a power section (51) acting as a drive, which comprises the stator (2), the rotary piston (4), the tooth chambers (53a, 53b) and the shaft external toothing (7) and is arranged, as viewed in the axial direction, between the housing section (10) on the output side and the inlet and outlet housing section (54),

   characterized in that

   • the secondary bearing (11) is arranged in the inlet and outlet housing section (54),

   • the disc-shaped rotary valve (12) is arranged, as viewed in the axial direction, between the power section (51) and the secondary bearing (11),

   • a cup-shaped sleeve (13) surrounds the shaft (6) in a radially spaced manner and tumbling about the shaft axis (52) and, as viewed in the axial direction, extends between the power section (51) and the inlet and outlet housing section (54) in the housing (24),

   • a first sleeve external toothing (14) of the sleeve (13) with a number of teeth Z5 on the side of the power section (51) engages the rotary piston internal toothing (5), and

   • a second sleeve external toothing (15) of the sleeve (13) with a number of teeth Z6 on the side of the inlet and outlet housing section (54) engages in a rotary valve inner toothing (16) with a number of teeth Z7 of the rotary valve (12) for the tumbling 1:1 rotary coupling between the rotary piston (4) and the disc-shaped rotary valve (12).
2. Hydrostatic gearwheel rotary piston machine according to claim 1, characterized by a difference in the number of teeth 2 between the number of teeth Z2 and the number of teeth Z1.
3. Hydrostatic gearwheel rotary piston machine according to claim 2, characterized in that

   • the number of teeth Z3 to the number of teeth Z4 has the values 10:11 or 11:12 or 12:13, and/or

   • the number of teeth Z1 to the number of teeth Z2 has the values 15:17 or 16:18 or 17:19.
4. Hydrostatic gearwheel rotary piston machine according to claim 2, characterized in that

   • the number of teeth Z3 to the number of teeth Z4 has the value 11:12 and

   • the number of teeth Z1 to the number of teeth Z2 has the value 16:18.
5. Hydrostatic gearwheel rotary piston machine according to one of claims 1 to 4, characterized in that

   • the number of teeth Z5 is equal to the number of teeth Z2 and

   • the number of teeth Z6 is equal to the number of teeth Z7.
6. Hydrostatic gearwheel rotary piston machine according to one of claims 1 to 5, characterized in that the hydrostatic axial balance piston (17) is arranged, as viewed in the axial direction, between the disc-shaped rotary valve (12) and the secondary bearing (11) in the inlet and outlet housing section (54).
7. Hydrostatic gearwheel rotary piston machine according to one of claims 1 to 6, characterized in that

   • as viewed in the axial direction, between the power section (51) and the inlet and outlet housing section (54), a control plate (25) with windows (26) and control plate channels (57) is arranged between the tooth chambers (53a, 53b) and the disc-shaped rotary valve (12) for supplying and discharging the working fluid to and from the tooth chambers (53a, 53b), and

   • the thickness (d) in the axial direction of the control plate (25) and/or its material are such that the control plate (25) is compliant such that the axial running play of the rotary piston (4) between the housing section (10) on the output side and the inlet and outlet housing section (54) reduces or remains the same with increasing working pressure of the working fluid.
8. Hydrostatic gearwheel rotary piston machine according to one of claims 1 to 7, characterized in that the shaft output side (8) of the shaft (6) is formed as a cone (37), for fastening a wheel flange (18) by means of an axial nut (19) to form a compact wheel motor.
9. Hydrostatic gearwheel rotary piston machine according to one of claims 1 to 8, characterized in that the stator internal toothing (1) is designed as rotatable rollers.
10. Hydrostatic gearwheel rotary piston machine according to one of claims 1 to 9, characterized in that the shaft external toothing (7) of the shaft (6) is conically shaped with a smaller diameter amount, as viewed in the axial direction, on its shaft output side (8).
11. Hydrostatic gearwheel rotary piston machine according to one of claims 1 to 10, characterized in that a tooth tip shortening is provided at the tooth tip of the shaft external toothing (7), in that the tooth flank radius (30) is provided mathematically larger at the points (28) and (29) of the tooth tip at which the toothings come out of engagement than at the point of deepest tooth engagement (31).
12. Hydrostatic gearwheel rotary piston machine,
    characterized by

    • a first hydrostatic gearwheel rotary piston machine (40) according to one of claims 1 to 11, and

    • a second hydrostatic gearwheel rotary piston machine (41) according to one of claims 1 to 11, wherein

    • the shaft opposite side (55) of the shaft (6) of the first hydrostatic gearwheel rotary piston machine (40) is coupled axially, in particular by means of a sleeve (42), to the shaft opposite side (55) of the shaft (6) of the second hydrostatic gearwheel rotary piston machine (41) and is connected in a torque-effective manner, and

    • the first hydrostatic gearwheel rotary piston machine (40) and the second hydrostatic gearwheel rotary piston machine (41) have, in particular, different intake quantities, in particular differently dimensioned tooth chambers (53a, 53b).
13. Hydrostatic gearwheel rotary piston machine according to claim 12, characterized by

    • a sleeve (2), by means of which the shaft opposite side (55) of the shaft (6) of the first hydrostatic gearwheel rotary piston machine (40) is axially coupled to the shaft opposite side (55) of the shaft (6) of the second hydrostatic gearwheel rotary piston machine (41) and connected in a torque-effective manner,

    • a first toothed shaft-toothed hub toothing (43a), which is connected to the sleeve (42) in a torque-effective manner, on the shaft opposite side (55) of the shaft (6) of the first hydrostatic gearwheel rotary piston machine (40), and

    • a second toothed shaft-toothed hub toothing (43b) of the second hydrostatic gearwheel rotary piston machine (41), which toothing (43b) is connected to the sleeve (42) in a torque-effective manner.
14. Hydrostatic gearwheel rotary piston machine according to claim 12 or 13, characterized in that

    • the sleeve (42) is radially mounted by a common rolling bearing (44) in the inlet and outlet housing section (54) of the first hydrostatic gearwheel rotary piston machine (40) and in the inlet and outlet housing section (54) of the second hydrostatic gearwheel rotary piston machine (41), and

    • the rolling bearing (44) is formed as a centering of the first hydrostatic gearwheel rotary piston machine (40) with the second hydrostatic gearwheel rotary piston machine (41).

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