EP0422617B1 - Suction regulated annular gear pump - Google Patents

Description

The invention relates to a suction-controlled gerotor pump with the features of the preamble of claim 1. The pump is generally driven by the shaft carrying the pinion. Such pumps are e.g. used to power hydraulic systems. In particular, the invention relates to the use of such a pump according to claim 11.

Motor vehicle engines and gearboxes in particular are operated in a large speed range. The speed parameters could behave like 10: 1 and above.

In contrast, the delivery target of the lubrication pump of a motor vehicle engine, which in automatic transmissions also has to take over the function of supplying pressure to the hydraulic switching elements and filling the converter against cavitation, is only approximately proportional to the speed in both the engine and the transmission in the lower third of the operating range. In the upper speed range, the oil requirement increases far less than the engine speed. It would therefore be necessary to have a drive-controlled lubrication or hydraulic pump or one with a speed-adjustable delivery rate. The most common form The oil and / or lubrication pump is the gear pump because it is simple, cheap and reliable.

The disadvantage is that the delivery rate (per revolution) cannot be regulated, i.e. the theoretical delivery rate is proportional to the speed. The practical characteristics of the delivery rate over the speed depend on a multitude of parameters such as delivery pressure, oil viscosity, flow resistance in the suction and pressure line, configuration of the toothing of the gearwheels, width of the gearwheels and design of the pump. An adjustment of the delivery line to the demand line, for example of an internal combustion engine, is in most cases too complex, which is why a bypass valve is used, which regulates the excess oil at a certain set delivery pressure and returns it decompressed to the suction line. This regulation is thus heavily lossy in the regulating range, so that the efficiency drops undesirably with increasing speed. The only practical way to avoid this excess quantity from a certain pump speed is the suction control. Since the flow resistances increase disproportionately with increasing oil speed, the static pressure in the suction opening of the gear chamber drops more and more until the so-called cavitation pressure threshold is reached, i.e. until the vapor pressure of the oil falls below The cell contents then consist partly of liquid oil, partly of oil vapor, partly also of sucked-in air, whereby it is under a static pressure that is significantly below atmospheric pressure. It is not a problem, e.g. by means of appropriately narrow suction lines or through an orifice or also controllable by means of a suction slide, the flow resistances in the intake manifold are to be determined or controlled in such a way that the useful delivery rate of the gear pump is largely adapted to the consumption line.

The cavitation that occurs is a disadvantage of this regulation. If the cell content, which is under low absolute pressure and partly consists of liquid and partly of gas, is suddenly transferred to zones of higher pressure, as is the case with such pumps in the case of such systems, then the gaseous ones implode Components of the cell contents so violent that undesirable noises and, what is worse, destruction of the cell walls are the result.

If a volumetric pump of this type is to be regulated by throttling on the suction side, then these implosions must be avoided. The procedure is known in such a way that on the displacement side of the pump, that is to say in the area of the shrinking cells, the cell contents have sufficient time to increase the static pressure to a sufficient extent by gradual compression so that at the moment , in which the cell communicates with the outlet channel, implosions of gas bubbles can no longer take place because these have already condensed to liquid again due to the constant reduction in the cell volume or have dissolved in the liquid (eg air). In terms of design, this solution can be solved most compactly with an internal gear pump in which the individual delivery cells are sealed off from one another. The time period for the slow compression of the steam and air spaces is structurally ensured by the fact that the cells on the displacement side of the pump are initially only connected to the delivery pressure chamber via check valves, so that the delivery pressure does not become effective in the case of a cell that is not completely filled with liquid can.

However, if the cells are already completely filled with liquid on the suction side, which, as explained at the beginning, is the case in the lower speed range, then the higher pinch pressure in the cell opens the check valve in the direction of the pressure delivery chamber, so that the displaced oil with only a slightly increased cell pressure can flow into the pressure chamber in relation to the delivery pressure corresponding to the opening pressure of the check valve and its flow resistance. Such a construction is known from DE-PS 30 05 657. This extends over the entire pressure half of the pump in the housing to the outlet channel leading axial bores that contain check valves at a distance from the gear chamber, which open only when the pressure of the cell in front of the respective bore exceeds the pressure in the outlet channel. Accordingly, this pump has one large axial extension. The spring valves used can break. The discontinuous connection of the feed cells to the outlet channel is also disadvantageous. Finally, the pressure distribution is also disadvantageous with regard to the use of cavitation-related implosions.

From the Patent Abstract of Japan, Vol. 10, No. 253 (M512) (2309), August 29, 1986, JP-A-6181588, a gerotor pump is known in which the outer gear running inside an internal gear has teeth with continuous circumferential teeth Has channels. These channels are intended to prevent peak pressures from occurring within the pumping chambers. However, this arrangement reduces the pumping capacity of the known internal gear pump, since a flow between the individual pumping chambers which compensates for the excess pressure is made possible.

From the Patent Abstract of Japan, Vol. 10, No. 333 (M534) (2389) dated November 12, 1986, JP-A-61138893, an internal rotor gear pump is known, in which grooves are provided in the area of the teeth of the internal gear are, which should allow an exchange of the liquid to be pumped with a sealed interior within the pump housing. In this way, on the one hand, a backward flow of the liquid to be pumped is to be prevented and, on the other hand, the noise development of the pump in connection with pressure fluctuations occurring is to be reduced.

The invention thus relates to a suction-controlled gerotor pump according to the preamble of claim 1, in which the difference in the number of teeth is 1 and the tooth shape of which ensures that the delivery cells are sealed from one another.

In particular, the invention solves the problem of creating a pump with a short and small diameter, which is also characterized by a favorable pressure curve in the pressure range, can also be retrofitted in existing constructions as a replacement for the lubrication pump, is reliable in operation and has a simple construction having.

This object is achieved by the characterizing features of claim 1.

The invention makes it possible, in most cases, to completely omit the bypass arrangement with a large passage, or to replace it with a small pressure relief valve, by adapting the conveying characteristic to the demand characteristic.

In the embodiment according to the invention, the housing is of extremely simple design and has only a very small axial extent. Because each feed cell can release working fluid into the leading feed cell when the feed cell shrinks when the ball valve is opened, but not in the opposite direction, the pressure in each feed cell in the reduction area can only be increased steadily until the pressure reaches the value has grown in the outlet opening. In this way, the feared implosions are avoided and the cavitation cavities are steadily reduced to zero. It is particularly advantageous here that a through the channels with the ball valves is not there is negligible flow resistance between the neighboring feed cells.

The arrangement of check valves in the teeth of the wheels is known per se from US-PS 35 15 496.

Basically, in the invention, for example, the mouths of the inlet and outlet channels can have recesses in the peripheral surface of the tooth chamber that supports the ring gear, the connection between the cells and the channel mouths then being effected by radial bores in the ring gear. However, the mouths of the inlet and outlet channels in the end walls of the gear chamber are preferably arranged as so-called inlet and outlet kidneys (claim 2). This allows very large inflow and outflow cross sections in and out of the feed cells.

The overflow channels can be provided, for example, in the gear bodies themselves. However, they are preferably arranged in the teeth of the wheels.

The check valves can e.g. be formed by cylindrical rollers arranged in corresponding widenings of the overflow channels and having an axis parallel to the pump axis, which under the influence of the flow lie in the widening against the corresponding channel mouth to be closed. Spring-loaded valves can also be used. However, the check valves are preferably designed as ball valves, the ball always striving to press the ball onto the valve seat due to the centrifugal force of the rotary movement of the gearwheel containing the valves. This training is not only simple in construction but also easier to manufacture and does not require valve springs.

In principle, the overflow channels can be designed, for example, as grooves in an end face of the corresponding gearwheel, a widening of the groove then accommodating the check valve. In this case, part of the wall of the overflow channels is formed by the corresponding end wall of the housing. So far there are several Opportunities. According to a preferred embodiment of the invention, however, the gear wheel containing the check valves is formed from two halves (the parting plane of which is a normal plane to the axis of rotation of the gear wheel), each of which contains half of the valve channels and the valve seat in mirror-image form.

The two halves do not necessarily have to be connected to one another, since they are fixed in their rotational position by the teeth of the corresponding gear and cannot move axially from one another, since this prevents the end walls of the gear chamber.

It should be taken into account here that the gear pump according to the invention with the number of teeth difference 1 is one in which all teeth are constantly in engagement with teeth of the counter gear. This ensures particularly good guidance of the two gear wheel halves against one another in the circumferential direction. The same also applies to centering.

However, it is preferred that the two halves of the wheel containing the overflow channels and check valves are connected to each other. The connection can be effected, for example, by explosion welding. Of course, the valve bodies must be inserted into the corresponding chambers before the weld connection.

Another possibility is that the two halves of the wheel are connected to one another by sintering. Finally, the two halves of the gearwheel containing the overflow channels can also be connected to one another by means of axial screws.

The two ring gear halves can be conventionally e.g. machined from appropriate blanks. According to a preferred embodiment of the invention, however, the two ring gear halves are produced in a powder-metallurgical sintering process. This allows you to do without any rework.

As a material for the gears come in the invention, for example, high-strength Sintered metals in question; however, depending on the intended use and the required number of pieces, steel or gray cast iron are also suitable as the material.

The valve body - preferably balls - can be steel balls, for example. However, balls made of non-metallic material or metal balls which are coated with a non-metallic material are preferably used here. This counteracts caking of the balls on the valve seats. The production from non-metallic material also reduces the inertial forces.

According to a preferred embodiment, the overflow channels are arranged in the teeth of the pinion and in this case have a cavity which accommodates the balls from one of the axial end faces of the pinion, the inflow and outflow channels to these cavities then being drilled.

A particularly good guidance of the valve balls is obtained if a support edge is provided in the non-return valve, which produces a tangential component of the centrifugal force on the ball in the direction of the valve seat. This allows the overflow channels to be guided in a particularly streamlined manner.

The preferred field of application of the invention is the use of the pump as an oil and / or hydraulic pump for motor vehicle engines and / or transmissions, in particular automatic transmissions. However, the invention is also applicable to other applications e.g. suitable in hydraulic control systems.

Further advantages of the invention result from the following description of preferred embodiments with reference to the attached schematic drawings.

• Fig. 1 eine vollständige Zahnringpumpe nach der Erfindung teilweise im Schnitt in einer Normalebene zu den Achsen der Zahnräder. (Hierbei sind die Rückschlagventile im Hohlrad angeordnet. Der Schnitt liegt in der Hohlradmitte),
• Fig. 2 einen vergrößerten Teilschnitt entlang der Linie A - A durch einen Hohlradzahn nach Fig. 1,
• Fig. 3 eine Teilansicht eines erfindungsgemäßen Zahnradsatzes, bei dem die Überströmkanäle im Ritzel angeordnet sind und der Schnitt ebenfalls etwa durch die Mitte des Zahnrades verläuft,
• Fig. 4 einen Schnitt durch einen Zahn des Ritzels gemäß Fig. 3 entlang der Linie B - B,
• Fig. 5 eine Teilansicht einer weiteren Ausführungsform der Erfindung, bei welcher der Schnitt durch das Hohlrad wieder in einer Normalebene zur Achse durch die Mitte des Hohlrades verläuft, und
• Fig. 6 einen Teilschnitt durch Fig. 5 entlang der Linie C - C.
• Fig. 7 zeigt schließlich die gemessenen Kennlinien einer Zahnringpumpe gemäß Fig. 1 und 2.

In these shows

• Fig. 1 shows a complete gerotor pump according to the invention, partly in section in a normal plane to the axes of the gears. (Here the check valves are arranged in the ring gear. The cut lies in the center of the ring gear),
• 2 shows an enlarged partial section along the line A - A through a ring gear tooth according to FIG. 1,
• 3 is a partial view of a gear set according to the invention, in which the overflow channels are arranged in the pinion and the section likewise runs approximately through the center of the gear,
• 4 shows a section through a tooth of the pinion according to FIG. 3 along the line BB;
• Fig. 5 is a partial view of another embodiment of the invention, in which the section through the ring gear again in a normal plane to the axis through the center of the ring gear, and
• 6 shows a partial section through FIG. 5 along the line C - C.
• 7 finally shows the measured characteristic curves of a gerotor pump according to FIGS. 1 and 2.

The pump shown in FIG. 1 has a pump housing 1 shown in simplified form, in the cylindrical gear chamber of which the ring gear 2 is mounted with its circumference on the peripheral wall of the gear chamber. The shaft 3 carrying the pinion 4 of the gerotor pump is also mounted in the pump housing. In this respect, however, other positions are also possible. The pinion has one tooth less than the ring gear, so that all teeth of the pinion are constantly in engagement with a tooth of the ring gear, as a result of which all the feed cells 13 and 17 formed by the tooth gaps of the pinion and ring gear are constantly sealed against the adjacent cells. The direction of rotation of the pump is clockwise, as indicated by arrow 18. In the end wall of the gear chamber lying behind the plane of the drawing in FIG. 1, the suction opening 11 is provided, which is dashed in the drawing is shown. The outlet opening 19 is also shown in dashed lines in the top left half. Intake and outlet opening are designed as so-called "kidneys".

The center points 5 and 6 of the gear wheels 2 and 4 have the center distance or the eccentricity 7, which together with the tip circle diameters of the gear wheels is responsible for the geometrically specific delivery volume of the running set. This is still proportional to the width 8 of the gears. These geometric variables determine the slope of the theoretical delivery line 9 of the pump shown in dashed lines in FIG. 7. At low speed, the suction speed in the inlet channel, not shown here, is low, so that in the suction kidney 10, which extends over almost the entire circumference of the suction area and is arranged laterally in the housing, the outline of which is shown by the broken line 11, the oil can flow in without bubbles, since none significant negative pressure occurs. The course of the negative pressure is shown at 12 below in FIG. 7. Since at this low speed and tooth frequency the flow impedance between the tooth and the tooth gap is small, the suction cells in the positions 13 between the teeth 14 and 15 in engagement are filled with largely bubble-free oil. As can be seen from the drawing, the mouth of the inlet channel or the suction kidney 10 extends in the circumferential direction up to close to the point 16, which is diametrically opposite the point of deepest tooth engagement. In the area of this point 16, the delivery cells formed by two tooth gaps opposite each other have reached their greatest volume and are completely filled with oil at low speed. If the pump then continues to rotate and the delivery cells reach the area to the left of point 16 in FIG. 1, the cells in positions 17 become displacement cells, since the volume of the delivery cells continuously increases from here to the point of deepest tooth engagement to almost zero decreased.

In non-regulated gear pumps of this type, the outlet opening 19, the outline of which is shown by the broken line 20, also becomes up to the point 16, and as far as possible, but not so far that a substantial leakage-effective short circuit can occur between the suction and pressure chamber. This means that the delivery cells in positions 17 can release the oil into the pressure channel at the start of their volume reduction without crushing losses. The outlet opening and thus the delivery cell in the first position 17.1 is under full delivery pressure. In contrast to this, in the embodiment of the pump according to the invention, the outlet opening of the gear chamber or the pressure kidney is shortened very far in the circumferential direction to the point of deepest tooth engagement, as can also be seen in FIG. 1. The delivery cells must also be able to empty themselves accordingly in positions 17.1 to 17.3 with bubble-free oil filling. This is made possible by the overflow channels 128 in the teeth of the ring gear 2. Each overflow channel 128 is provided with a check valve 21. It can be seen that the delivery cells in positions 17.1 to 17.3, in which their volume is steadily decreasing, can empty through the series-connected overflow channels 128 with the check valves 21.1 to 21.3 arranged in them in the delivery direction towards the pressure kidney. In this case, a somewhat higher static pressure must then prevail in the delivery cells in positions 17.1 to 17.3 than in the outlet opening of the pressure kidney 19, since the overflow channels 128 with the check valves 21 are naturally lossy with respect to the flow resistance. At low speed these losses are not high because the flow velocities are low. These throttling losses should of course be kept as small as possible by means of a corresponding design of the check valves.

The mouths of the overflow channels and / or the tooth and tooth gap shape must of course lie or be dimensioned such that a liquid flow in the direction of pump rotation is prevented at the point of deepest tooth engagement. This is not a problem.

Up to a certain limit speed, a delivery rate that is in principle proportional to the speed is also delivered in the pump according to the invention. If this limit speed is exceeded, so the static pressure in the feed line begins to drop and drops below a critical value, as can best be seen in FIG. 7. In the pump examined, this speed range is around 1200 rpm. From 1450 rpm the flow rate stagnates despite the increasing speed, since the static suction pressure has fallen below the evaporation pressure of the oil. From now on, cavities are created in the delivery cells in positions 13, which theoretically concentrate in the area of the root circle 22 of the pinion 4, since the bubble-free oil is forced radially outwards by centrifugal force. At about 2100 rpm, the pump delivers only 2/3 of its maximum delivery volume, as can be seen from FIG. 7. This state is shown in Fig. 1 by a dashed level line 23 as a concentric circle to the center of the ring gear. This level line 23 is provided with the level symbol 24. Radially inside the level line there is essentially oil vapor and / or air, radially outside there is essentially oil. The level line 23 passes through the tooth base point 25 of the delivery cell in position 17.3, which is in the process of being connected to the pressure kidney or outlet opening 19. The pump is advantageously designed such that, even at the maximum operating speeds to be expected, the level line does not move radially outward much further than to the base of the pinion tooth gap of the delivery cell, which is just beginning to reach the edge of the outlet opening 19.

This level line can of course always be located radially further inside, as long as the suction control does not suffer.

Since the delivery cells in positions 17.1 to 17.3 are sealed against each other by tooth flanks or tooth tip engagement and the check valves in the construction shown are not only due to the centrifugal force acting on the valve ball on the one hand, but also due to the static increase from cell positions 17.1 to 17.2 to 17.3 Pressure are closed, the delivery pressure in the outlet opening 19 cannot act into the delivery cells in positions 17.1 to 17.3. The cavities 26 within the leveling ring surface 23 thus have enough time to reduce until the position 17.3 is reached by reducing the cell volume until finally the Cell in position 17.3 connects to the pressure line. The feared cavitation caused by sudden imploding of the cavities is thus avoided.

As can be seen from the position of the level line 23 in FIG. 1, cavitation can actually be expected again at speeds above 2100 rpm, since the degree of filling of the pump then drops further, as shown in FIG. 7. In practice, however, it had been shown that the transition here is very grinding and that cavitation noises could not be heard even at a significantly higher speed. This should be caused by the fact that dynamic influences continue to cause a very gentle pressure increase from the feed cell position 17.1 to position 17.3.

In FIG. 2, a section through the centrifugal ball check valve arrangement from FIG. 1 is shown in a greatly enlarged illustration. The ring gear here consists of two halves which are soldered or welded to one another in the parting plane indicated by the parting lines 27 and 28. To the left and right of the ball 29, 30 bypass channels 30 are provided so that when the valve seat 31 is open, there is sufficient passage cross section.

In the embodiment shown in FIGS. 3 and 4, the overflow channels 33, 34 are generated in the teeth of the pinion by drilling. The sprocket made here, for example, of steel is undivided. To form the check valve, a cavity 35 is incorporated into the teeth from one end face of the pinion, which has a supporting edge 32 which, like the construction to be described later, according to FIGS. 4 and 5, serves to guide the ball 36 during the closing movement . If the cavern is not sintered, which is the cheapest, it can also be milled using an NC-controlled milling machine, for example. The overflow channels 33 and 34 can be simply drilled here. The balls 36 are also automatically pressed centered on the valve seat by the centrifugal force and the hydrostatic force. They are prevented from falling out by the housing wall 37.

As can be seen from the drawings, the channels with the ball valves should always be guided in such a way that the centrifugal force already tries to press the valve balls onto their seats. In other words, in the preferred embodiment, the valve channels should be curved in such a way that the ball movement, as is the case in FIG. 1, has an essential radial component. If you do not have such a possibility, you can use a support edge 32, around which the ball can tilt, so that the ball is first pressed against the support edge 32 by the centrifugal force and continues under the influence of the centrifugal force around this edge 32 into its Valve seat closing position can pivot.

In the embodiment shown in FIGS. 5 and 6, the overflow channels and the check valves are arranged in the ring gear, but are somewhat more streamlined than in the embodiment according to FIGS. 1 and 2. For this purpose, a support edge 32 is provided, which is caused by the centrifugal force Tangential closing force component generated so that the valve seat has a tangential line of action C - C. Such a design is recommended when the gear set has to be made very wide. In this case, a lot more oil must flow through the check valves at low speed and without throttling.

1 and 2 and 5 and 6 can be made possible by axial division of the gearwheels, the gearwheel halves being able to be produced using the powder metallurgical process. Since the fatigue strength of such powder-metallurgically manufactured components is limited, the pressure output of the pump is also limited in this case.

If you want to avoid this disadvantage of powder metallurgical production, the pump can be manufactured, for example, according to FIGS. 3 and 4.

Claims (12)

1. A suction-controlled gear ring pump comprising

   - a housing,

   - an internally geared hollow gear (2) rotatably arranged in a gear box of the housing (1),

   - a pinion (4) having one tooth less than said hollow gear (2) and engaging with and arranged in said hollow gear (2), the teeth of said pinion forming, together with the teeth of said hollow gear (2) alternately expanding (13) and reducing (17) successive feed cells for the operating liquid and providing sealing between said feed cells,

   - inlet and outlet passages arranged in the housing (1) for the entry and discharge of the operating liquid, said passages opening out (10, 19) into the gear box on either side of the location of deepest tooth engagement,

   - a fixed or variable throttle provided in the inlet passage,

   - and check valves (21) in the pressure region of the pump,

   characterized in

   - that the end of the mouth (19) of the discharge passage remote from the location of deepest tooth engagement is positioned so close to said location of deepest tooth engagement that several feed cells (17) are present at all times between said mouth end and the circumferential location (16) where said feed cells (13, 17) are beginning to diminish,

   - that said feed cells (13, 17) are respectively connected to the neighbouring feed cells by overflow channels (128) provided in at least one of said gears (2, 4),

   - and that the check valves (21) are positioned in such a way in the overflow channels (128) that they counteract a flow of operating liquid against the feed direction.
2. A gear ring pump according to claim 1, characterized in that the mouths (10, 19) of the inlet and discharge passages are positioned in the front walls or one front wall, respectively, of the gear box.
3. A gear ring pump according to claim 1 or 2, characterized in that the overflow channels (128) are positoned in the teeth of the gears (2, 4).
4. A gear ring pump according to one of the claims 1 to 3, characterized in that the check valves (21) are formed as ball valves, the centrifugal force of the rotating motion of the gear containing the valves aiming to press the ball onto the valve seat.
5. A gear ring pump according to one of the claims 1 to 4, characterized in that the gear (2) containing the check valves consists of two halves which, in mirror reversed form, respectively contain one half of the overflow channels (128) and of the valve seat.
6. A gear ring pump according to claim 5, characterized in that the two halves of the gear have been produced by a powder metallurgy method.
7. A gear ring pump according to claim 5, characterized in that the two halves of the gear have been joined by explosion welding.
8. A gear ring pump according to claim 5, characterized in that the two halves of the gear have been joined by sintering.
9. A gear ring pump according to one of the claims 4 to 8, characterized in that the valve balls (21) consist of non-metallic material or are coated with a non-metallic material.
10. A gear ring pump according to claims 1 to 3, characterized in that the overflow channels (33, 34) are positioned in the pinion and comprise recesses (35) worked into said pinion starting from the axial front wall and receiving the valve balls, said recesses comprising drilled inlet and discharge passages (33, 34).
11. A gear ring pump according to one of the claims 4 to 10, characterized in that a supporting edge (32) is provided in the check valve, said supporting edge generating on the ball a tangentially effective component of the centrifugal force towards the valve seat.
12. Use of the gear ring pump according to one of the claims 1 to 11 as oil and/or hydraulic pump for automobile engines and/or transmissions.

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