EP0712997A2 - Valve control with suction regulated internal gear pump - Google Patents

Abstract

A pump driven by the engine is used for supplying the valve timing and lift adjustment component with work fluid. The pump is formed as a suction-regulated gear ring pump with a sealing plate extending over several delivery cells (17.1-17.3) and has a speed-dependent delivery characteristic line which is matched to the work fluid requirement of the adjustment component. The valve control components are camshafts, the phase position of which is alterable for control of the overlapping times of inlet and exhaust valves. The pump supplies with work fluid an adjustment component for altering a valve lift. The pump also supplies the engine with lubricating oil which also acts as work oil for the hydraulic adjustment component.

Description

The invention relates to a valve control for an internal combustion engine according to the preamble of claim 1 and to a suction-controlled toothed ring / internal gear pump which can be used in particular for the latter according to the preambles of claims 16 and 28.

In the course of the progressive development in the automotive industry, the demands on the engine performance are constantly increasing. The engines should be optimally controlled over a wide speed range. To meet this requirement in both the bottom and To be able to meet in the upper speed range of the engine, valve controls have been developed with which the overlap times of intake and exhaust valves can be changed depending on the speed. In controls for adjusting the valve overlap times, known as so-called VTC (valve timing control) controls, the camshafts for each of the intake valves and the exhaust valves are adjusted relative to one another, so that the cams of the two camshafts experience a phase shift.

In addition to this camshaft adjustment by mutually rotating the camshafts, the valve strokes can also be changed. Large valve strokes with correspondingly longer overlap times in the upper speed range and smaller valve strokes with short or no overlap times in the lower speed range of the engine are set. Furthermore, an adjustment of the valve lift and / or the overlap times from warm-up operation to normal operation is desirable.

A multi-phase valve timing mechanism is known from "Motortechnische Zeitschrift" 55 (1994) 6, page 342. The cam set used for a six-cylinder engine has two rocker arms. T-shafts control the two intake and exhaust valves per cylinder simultaneously, depending on the speed. At high speeds, hydraulic pistons connect the corresponding rocker arms to the T-shafts. At low speed, the T-shafts are connected to the levers for low speeds. This mechanism also enables cylinder deactivation. For this purpose, the T-shafts are released from the rocker arms for the high speeds, so that only three of the six cylinders are still working.

Ordinary pumps for pumping motor oil, for example vane pumps or conventional gear pumps, pump their working medium with a constantly increasing delivery pressure or delivery volume flow with the pump speed. The Pumps are usually driven directly mechanically by the motor via a corresponding toothed belt drive or another suitable gear, so that the delivery pressure or volume flow increase with the motor speed. In order to be able to carry out the required valve control processes even at low engine speeds, the pumps which can be used must have a steep increase in their volume flow delivered in the lower speed range of the engine. The known pumps are therefore large with a correspondingly high power consumption. With increasing engine speed, they therefore pump more engine oil than is required by the valve control actuators, so that the excess must be returned directly from the pump outlet to a sump.

A pump designed as an internal gear pump is e.g. known from DE 39 33 978. The drive is usually carried out by the shaft carrying the pinion. The delivery target of such pumps, e.g. the lubrication pump of a motor vehicle engine is only approximately proportional to the speed in the lower part of the operating range. In the upper speed range, the lubricant or working fluid requirement increases far less than the speed of the engine. Suction control of the pump is therefore necessary.

The cavitation that occurs is disadvantageous in such a suction control. The linear pressure rise to be expected due to the increase in the speed cannot be maintained in the pressure range of such pumps; rather, the pressure does not rise linearly with a smaller rise from a certain speed. If the geometric delivery rate in the working area falls below the proportionality area, cavitation occurs, which leads to implosions of the gaseous components of the cell contents, so that undesirable noises and damage to the cell walls are the result. Furthermore, such pumps have relatively low efficiencies in higher speed ranges.

The invention has set itself the task of providing a valve control for an internal combustion engine in which actuators for adjusting control means for valves of the engine can be supplied with the working fluid necessary for actuating the actuators in an energy-saving and therefore inexpensive manner. It is a further object of the present invention to provide an internal gear pump with minimal cavitation and high efficiency, which can be used in particular for a valve control mentioned above.

These objects are solved by the subject matter of claims 1, 16 and 28.

Preferred embodiments are described by the subclaims.

A valve control for an internal combustion engine is equipped according to the invention with a suction-controlled gerotor pump which has a sealing web with a plurality of delivery cells, the so-called pressure cells, which decrease in size from an inlet for working fluid to a pump outlet. Such a pump used for the purposes of the invention already has a speed-dependent delivery characteristic which essentially corresponds to the requirements of the valve control. In its lower speed range, such a pump has a steep increase in the delivery rate in order to be able to supply all consumers with sufficient oil immediately. The delivery curve flattens out in the upper speed range or is essentially constant there, which corresponds to the actual need for a valve control. The hydraulic power loss can be reduced in this way. With a corresponding design of the pump, the expensive pressure control valves necessary according to the state of the art can be dispensed with. Simple safety valves are sufficient to protect particularly sensitive consumers from overpressure when the engine is cold started. By adapting the delivery rate to the demand, a hydrostatic power saving is achieved Savings achieved in the components in the pump delivery circuit.

A suction-controlled gerotor pump is advantageously used as a feed pump for adjusting the camshaft. Another preferred use is as a feed pump for valve lift adjustment. Furthermore, such a pump can advantageously be used for switching cylinders on and off, as described, for example, in the aforementioned "Motortechnische Zeitschrift" 55 (1994) 6, page 342. A combination of such valve control types can also be advantageously supplied by such a suction-controlled gerotor pump. With appropriate dimensioning, the pump used according to the invention for the purpose of valve control can additionally supply the engine with lubricating oil. The lubricating or engine oil also serves as working oil for the valve control actuators.

The pump preferably has a throttling on the suction side, which can be changed in order to be able to adapt the pumping characteristics of the pump even better to the needs of the consumers. In this way, a pump with a multi-stage delivery characteristic, the number of stages of which corresponds to that of the throttle, can be provided with a multi-stage throttling. Simple orifices or throttles can also be used as throttling elements. A continuous adjustability of the throttling can also be used to advantage in order to be able to flexibly adapt pumps of one pump size to different requirements on site.

The decisive advantage of the new internal gear pump according to the invention is that the controlled supply of working fluid from the outlet opening into an inlet opening and the simultaneous interruption of the supply of working fluid from the inlet duct into this inlet opening result in a delivery cell in which pressure drop and thus cavitation occur with increasing speed would be brought to the higher outlet pressure. This prevents cavitation in this delivery cell. Furthermore, a great advantage arises from the fact that, because there is no cavity, ie no negative pressure, in this feed cell, but instead it is pressurized, this pressure generates a positive torque on the pinion. This feed cell, which is under higher pressure, thus works like a hydraulic motor, which means that very high efficiency can be achieved.

According to a preferred embodiment, the device mentioned in the characterizing part of claim 16 connects the inlet orifices adjacent to it to the pressure area in succession with increasing pressure in the pressure area. As a result, as the speed increases, it is ensured that the respective delivery cell in which the pressure drop and thus cavitation could occur is pressurized in good time so that noise and damage can be avoided.

Advantageously, the above-mentioned device has a transfer duct connected to the outlet opening, which opens into at least one supply duct via a valve device, which in turn is connected to an inlet opening. The valve device can thus control the regulated supply of working fluid from the outlet opening, that is to say the pressure range, into the inlet opening and at the same time initially restrict the supply of working fluid from the inlet duct into this inlet opening and interrupt it later. For this purpose, such a valve device preferably has a valve piston which is mounted against the pressure of the working fluid in the transfer duct by means of a spring supported in the housing and which blocks or enables access of the working fluid into the supply ducts by means of a shoulder. With different stiffness selections, the spring provides a way to control the performance of the valve assembly while the head of the valve piston can be designed so that the pressurized working fluid presses against the spring force against one of its surfaces, while it blocks or releases with its side surfaces the supply channels for the working fluid depending on the position of the valve piston.

The valve piston can be held in the depressurized state of the transfer channel or up to a predetermined pressure in it against the force of the spring by a stop on the housing in a position where no working fluid flows from the transfer channel into a supply channel. This state corresponds to the initial position of the valve device at low speed or when the pump is at a standstill. The opposite stop point of the valve piston can be determined in that the valve piston is stopped in the position where working fluid flows from the transfer channel into all supply channels in its movement against the direction of the spring force because the spring is blocked.

The size of the inlet opening for the conveyor cells that are not to be connected to the transfer channel is preferably limited to the area in which these conveyor cells extend. This ensures that those feed cells that are to be pressurized with pressure from the high-pressure chamber with increasing speed can be completely cut off from the suction chamber. In contrast, the outlet opening can extend approximately over the entire area of the conveyor cells, which are downstream of the conveyor cells in the conveying direction and can be connected to the transfer channel. This design of the outlet mouth is suitable because the conveyor cells connected to it are under high pressure practically during the entire operation.

In a preferred embodiment, the end of the valve piston facing away from the head shoulder forms, together with the housing, a spring chamber which is used for Damping of the piston movement is filled with working fluid and is in fluid communication with the working fluid in the inlet channel via a bore.

The valve device advantageously acts simultaneously as a safety valve in the form of a bypass valve. If, at maximum pressure in the pressure area, the head heel has exceeded the last supply channel to such an extent that a short-circuit flow of the working fluid from the pressure area into the inlet channel occurs under decompression, the spring therefore only goes into block when a sufficient outflow cross section is created.

In a further advantageous embodiment of the present invention, the pinion of the internal gear pump has two teeth less than the toothed ring, and at the point of disengagement of the teeth, a crescent-shaped filler piece is provided. In this case, the teeth of the toothed ring should be designed to be sufficiently pointed so that the feed cells in the suction area are sealed against one another via the tooth engagement.

Furthermore, the internal gear pump according to the invention can be characterized in that the head shoulder of the valve piston consists of a heel base and a longitudinally adjoining shoulder tab with the same outer diameter, the guidance and the sealing function of the valve piston in the housing bore on the housing shoulders on the outer surfaces of the heel base and the Paragraph flag take place.

An internal gear pump according to the invention can advantageously be used as a suction-controlled pump for a valve control according to this invention.

Figur 1

eine Darstellung des Arbeitsölbedarfs einer Ventilsteuerung;

Figur 2

eine sauggeregelte Zahnringpumpe mit Drosselblende im Einlaßkanal;

Figur 3

die Förderkennlinie der sauggeregelten Zähnringpumpe nach Figur 2;

Figur 4

eine sauggeregelte Zahnringpumpe im Querschnitt;

Figur 5

eine weitere sauggeregelte Zahnringpumpe im Querschnitt;

Figur 6

den Leckölstrom als Funktion der Drehzahl N für die Pumpe nach Figur 5;

Figur 7

den Saugdruck im Einlaß der Pumpe nach Figur 5 als Funktion der Pumpendrehzahl;

Figur 8

den Zwischendruck PI und die Druckdifferenz PI-PH bei der Pumpe nach Figur 5 als Funktion der Pumpendrehzahl;

Figur 9

eine Querschnittsansicht einer erfindungsgemäßen Innenzahnradpumpe, bei der die Stellung der Ventileinrichtung im Anlaufzustand der Pumpe wiedergegeben ist;

Figur 10

eine Querschnittsansicht der erfindungsgemäßen Innenzahnradpumpe in einem Zustand mit gegenüber der Figur 9 erhöhter Drehzahl;

Figur. 11

eine Querschnittsansicht der erfindungsgemäßen Innenzahnradpumpe, wobei die Drehzahl soweit angestiegen ist, daß die Ventileinrichtung bereits eine von der Zufuhr durch ihre Einlaßmündung abgetrennte Förderzelle zur Druckbeaufschlagung vom Druckbereich her freigibt;

Figur 12

eine Querschnittsansicht der erfindungsgemäßen Innenzahnradpumpe, bei der die Ventilvorrichtung eine Stellung eingenommen hat, in welcher alle Einlaßmündungen und Zufuhrkanäle die mit ihnen verbundenen Förderzellen mit unter Hochdruck stehender Arbeitsflüssigkeit versorgen; und

Figur 13

eine weitere Ausführungsform der erfindungsgemäßen Innenzahnradpumpe, wobei das Ritzel zwei Zähne weniger aufweist als der Zahnring und an der Stelle des Außereingriffkommens der Zähne ein mondsichelförmiges, gehäusefestes Füllstück vorgesehen ist.

The invention is described below on the basis of exemplary embodiments shown in the figures explained in more detail. Show it:

Figure 1

a representation of the working oil requirement of a valve control;

Figure 2

a suction-controlled gerotor pump with throttle orifice in the inlet duct;

Figure 3

the delivery characteristic of the suction-controlled gear pump according to Figure 2;

Figure 4

a suction-controlled gerotor pump in cross section;

Figure 5

another suction-controlled gerotor pump in cross section;

Figure 6

the leakage oil flow as a function of the speed N for the pump of Figure 5;

Figure 7

the suction pressure in the inlet of the pump of Figure 5 as a function of the pump speed;

Figure 8

the intermediate pressure PI and the pressure difference PI-PH in the pump according to FIG. 5 as a function of the pump speed;

Figure 9

a cross-sectional view of an internal gear pump according to the invention, in which the position of the valve device is shown in the starting state of the pump;

Figure 10

a cross-sectional view of the internal gear pump according to the invention in a state with increased speed compared to Figure 9;

Figure. 11

a cross-sectional view of the internal gear pump according to the invention, wherein the speed has risen so far that the valve device already releases a feed cell separated from the supply through its inlet opening for pressurizing from the pressure area;

Figure 12

a cross-sectional view of the internal gear pump according to the invention, in which the valve device has assumed a position in which all inlet openings and supply channels supply the feed cells connected to them with working fluid under high pressure; and

Figure 13

a further embodiment of the internal gear pump according to the invention, wherein the pinion has two teeth less than that Toothed ring and a crescent-shaped, housing-fixed filler is provided at the point of disengagement of the teeth.

FIG. 1 shows a volume flow V _{P of} a pump and a volume flow requirement of a valve control as a function of the engine speed D _M. The volume flow requirement of the valve control initially increases up to an engine speed D1 _M , remains essentially constant in the subsequent speed range between D1 _M and D2 _M , increases a second time from the speed D2 _M to an engine speed D3 _M again thereafter to remain essentially at the value reached at D3 _M with further increasing engine speeds.

FIG. 2 shows a suction-controlled gerotor pump 100 which, due to the suction control, already has a delivery characteristic which is adapted to the volume flow requirement of a valve control. The delivery characteristic of the suction-controlled gerotor pump according to FIG. 2, namely its volume flow V _P plotted against the pump speed, which can also be thought of as being replaced by the pump delivery pressure, is shown in FIG. 3. Thereafter, the volume flow V _{P delivered} by the pump flattens or kinks from a limit speed D _g which can be determined by design or which can also be set during operation, the so-called cut-off point, and remains essentially constant thereafter, despite a further increasing pump speed D _P.

The amount of oil at the cut-off point D _{g is} limited by a throttle orifice 14 in the intake manifold or inlet channel 12 of the pump 100. A critical flow rate is established at the throttle orifice 14, and the amount of oil drawn in and delivered remains essentially constant despite a further speed increase from the cut-off point D _g . The throttling on the intake side results in a strong negative pressure after the orifice 14, which is lower than the vapor pressure of the oil. The oil begins to boil and evaporate. With a rotation of an internally toothed Ring gear 2 and an engaging pinion 4 above the cut-off point D _g fill the tooth chambers 13 with an oil-gas mixture via an inlet opening into the interior of the pump, the so-called suction kidney 11. In a conventional gerotor pump, the sealing web between the suction kidney 11 and a pump outlet, the so-called pressure kidney 20, is small. If such a pump were used, the tooth volume under low pressure would suddenly be pressurized. The "high pressure oil" would penetrate into the "low pressure area" and the gas bubbles would suddenly change from the gaseous state to the liquid state, ie they would implode. This phenomenon, known as "cavitation", causes noise and damage to the pump. To avoid this, the suction-controlled gerotor pump has a long sealing web between the suction kidney 11 and the pressure kidney 20. The sealing web should cover an angle of at least 45 °, preferably at least 90 °. The oil / gas mixture is compressed slowly and not suddenly at maximum tooth chamber volume and after suction and subsequent volume reduction by rotating the pump. The gas can pass through a controlled change of state in the pressure cells 17 forming the sealing web and have passed into the liquid state before the tooth chamber volume is emptied into the pressure kidney 20.

In the lower pump speed range before the cut-off point D _g , the tooth chambers 17 lying along the sealing web between the suction cardioid 11 and the pressure cardioid 20 are filled 100% with oil. Starting from the maximum tooth chamber volume, the suction kidney edge is overlapped when the wheel set 2, 4 is rotated, the tooth chamber volume is separated and, upon further rotation, pressure is applied by reducing the volume. Now ball valves 21 come into operation, which are arranged in the outer ring gear 2 in overflow channels 128 and act as check valves. If the pressure in a tooth chamber 17 rises, the trailing valve 21 is closed with respect to the suction kidney 11 acting as a suction space, the leading valve 21 is opened in relation to the pressure kidney 20, which acts as a pressure chamber. The oil flows into the next tooth chamber via the bypass channel formed in this way. Since the pressure is also increased there during rotation, the oil flows into the tooth chamber that then follows and so on until it reaches the pressure kidney 20. It could be demonstrated by measurement that this pump does not generate cavitation. The oil can form gas bubbles, but they do not implode, but slowly and controlledly change to the liquid state.

In the case of a gerotor pump which is throttled on the intake side to a regulating point D _g and which is designed as described above, the desired steep increase in the required oil volume flow V _P at a low pump speed can thus be achieved with a corresponding dimensioning of the pump. Despite the oil / gas mixture forming in the sealing web between the suction kidney 11 and the pressure kidney 20 as the pump speed D _P increases, the power consumption of the pump remains comparatively low at a substantially constant volume flow V _P. When using such a pump in the supply circuit of a valve control, little or no excess oil has to be drained into a sump. The use of expensive pressure control valves can also be dispensed with. At best, cheap pressure relief valves are necessary. Compared to the pumps conventionally used, the power saving corresponds approximately to the volume flow triangle above the cut-off point D _g , ie approximately to the upper triangular area shown in dark in FIG. 3.

FIG. 4 shows a pump which is particularly suitable for the purposes of the invention, as is known from DE 42 09 143 C1. This pump has a pump housing 1 shown in simplified form, in the cylindrical gear chamber of which the ring gear 2 is supported 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 1; however, other positions are also possible.

The pinion 4 has one tooth less than the ring gear 2, so that each tooth of the pinion 4 is constantly in engagement with a tooth of the ring gear 2, as a result of which all the feed cells formed by the tooth gaps of the pinion and ring gear are constantly sealed against the adjacent cells. The pump turns clockwise. The suction kidney 11 is provided in the end wall of the gear chamber located behind the plane of the drawing. The same applies to the pressure kidney 20. The center points of the two gear wheels 2 and 4 have an eccentricity which, together with the tip diameter and the width of the gear wheels, determines the steepness of the pump delivery line (FIG. 3).

At low speed, the suction speed in the intake manifold 12 is low, so that the oil can flow in bubble-free from the suction kidney 11, which extends almost over the entire intake circumference area and is arranged laterally in the housing 1, since no significant negative pressure occurs. Since the flow impedance between tooth and tooth gap is also small at low speed and tooth frequency, the suction cells 13 formed by the teeth of the wheels 2 and 4 of the suction side are filled with largely bubble-free oil. The suction kidney 11 serving as the intake pipe mouth extends in the circumferential direction of the wheels 2 and 4 up to close to a point 16 of minimal tooth engagement. In the area of this point 16, the delivery cells 13 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, the cells in positions 17.1, 17.2 and 17.3 become displacement cells, since the volume of the delivery cells extends from there to point 7 of the deepest meshing, which is the point 16 the smallest tooth engagement is diametrically opposed, continuously reduced to almost zero.

In the case of non-suction-controlled gerotor pumps, the pressure kidney 20 serving as the outlet opening can reach close to the point 16. This would make the pressure kidney 20 and thus the delivery cell in the first position 17.1 are under full delivery pressure.

In contrast, in the present pump, the pressure kidney 20 of the gear chamber is shortened very far in the circumferential direction to the point of deepest tooth engagement, so that a plurality of delivery cells 17.1 to 17.3 lie between the suction kidney 11 and the pressure kidney 20. In the exemplary embodiment, the sealing web covers an angle of more than 90 °. The pumping cells 17.1 to 17.3 must be able to empty with bubble-free oil filling. This enables overflow channels 128 in the teeth of the ring gear 2. Each overflow channel 128 is provided with a check valve 21. The delivery cells 17.1 to 17.3, in which the volume of the compressed medium is constantly decreasing, can empty through the series-connected overflow channels 128 with the check valves 21.1 to 21.3 arranged therein in the direction of delivery to the pressure kidney 20. In this case, a somewhat higher static pressure must prevail in the delivery cells 17.1 to 17.3 than in the pressure kidney 20, since the overflow channels 128 with the check valves 21 are lossy because of the flow resistance. At low speed these losses are not high because the flow velocities are low. The throttle losses should be kept as small as possible by means of an appropriate design of the check valves.

Up to a certain limit speed D _g (FIG. 3), an approximately proper flow rate is delivered. If this limit speed D _{g is} exceeded, the static pressure in the intake manifold 12 begins to drop and drops below a critical value. In the pump examined according to the exemplary embodiment, this limit speed D _{g is} approximately 1200 rpm. From around 1500 rpm, the flow rate stagnates despite the increasing speed, since the static suction pressure has fallen below the evaporation pressure of the working oil. From now on, voids are created in the feed cells of the suction side of the pump, which are theoretically in the Concentrate the root circle of the pinion 4, ie at 22, since the bubble-free oil is forced radially outwards by centrifugal force. At around 2100 rpm, the pump only pumps around 2/3 of its maximum delivery volume. This state is represented 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 23 there is essentially oil vapor and / or air radially outside essentially oil. The level line 23 passes through the base point 25 of the pinion tooth gap of the delivery cell 17.3, which is in the process of being connected to the pressure kidney 20. The pump is advantageously designed such that, even at the maximum operating speeds to be expected, the level line 23 does not move radially outward much further than to the base point 25 of the pinion tooth gap of the delivery cell 17.3, which is just beginning to reach the edge of the pressure kidney 20. Of course, radially further inward, this level line 23 can always lie as long as the suction control does not suffer.

Since the feed cells 17.1 to 17.3 are sealed against each other by tooth flank or tooth tip engagement and the check valves 21 in the construction shown not only by the centrifugal force acting on the valve ball on the one hand, but also by the static pressure rising from the cells 17.1 via 17.2 to 17.3 are closed, the delivery pressure in the pressure kidney 20 cannot act in the delivery cells 17.1 to 17.3. The cavities within the leveling ring surface 23 therefore have enough time to break down until the delivery cell 17.3 is reached due to the volume reduction of the cells.

In order to shift the limit speed D _g upward, a bypass is provided in the intake manifold 12 parallel to the orifice 14, in which a further throttle, namely a throttle valve 43, is arranged, which can be adjusted between the "open" and "lock" positions .

The pump designed in this way with the throttle diaphragm 14 and the throttle valve 43 arranged parallel thereto is already adapted to the demand curve of the valve control according to FIG. 1. It is only necessary to change the throttle valve 43 from its "lock" position to the "open" position at the engine speed D2 _{M shown} in FIG. 3.

Furthermore, the outlet channel 19 of the pressure kidney 20 is not only fed from the pressure kidney 20, but also from a further outlet opening 35 connected upstream of this pressure kidney 20, which is connected via a channel 36 to the outlet channel 19 in the manner shown in FIG. In the channel 36 there is also a throttle valve 37 which is adjustable or switchable between a position blocking the channel 36 and a position which enables the flow through the channel 36.

In the normal operating state, the two throttle valves 43 and 37 are closed. If larger quantities of oil are now required because an adjusting means 76 or 82 is switched on, a corresponding control device opens the two throttle valves 43 and 37. On the one hand, this greatly reduces the intake resistance and accordingly moves the level line 23 outwards. In FIG. 3, the limit speed D _{g of} the conveying characteristic curve moves upwards along the oblique line. The opening of the throttle valve 43 is coupled to the pump speed and thus to the motor speed via suitable control electronics so that the valve 43 is opened, for example, when the motor speed D2 _M entered in FIG. 3 is reached.

Because the throttle valve 37 is also switched when the throttle valve 43 is switched over, the now larger amount of oil does not have to be displaced additionally through the overflow channels 128 to the front end of the pressure kidney 20. Due to the upstream outlet opening 35 and the channel 36, the functionally decisive edge of the pressure kidney 20 is now closer to the point 16 least meshing. In this way, throttling losses in the overflow channels 128 are minimized. The efficiency of the pump is increased and the delivery rate increases approximately linearly until the speed of the motor has reached the new, higher limit speed.

Other throttle arrangements in the intake manifold 12 are possible. If a bypass is omitted, the arrangement of a single step-by-step or continuously adjustable throttle valve can also be used advantageously. A control valve can also be provided. The throttling in the intake manifold 12 - and also in the exhaust duct 19, 36 - is controlled as a function of the engine speed, on which the working oil requirement of the valve control of the engine also depends. The suction-controlled gerotor pump can thus be adapted to a wide variety of demand lines by appropriate throttle arrangements.

In addition to the overflow channels 128 equipped with non-return valves 21, an additional bypass can be provided in the end wall of the gear chamber in the way of the delivery cells 17.1 to 17.3, namely near the tooth root circle of the ring gear 2, which extends in the circumferential direction to the front edge of the pressure kidney 20 . The formation of such a bypass is known from the application P 43 30 586.5 and is shown in FIG. 5.

In accordance with the relatively large number of teeth, this bypass is formed by openings formed in the end wall of the gear chamber, in the exemplary embodiment there are two openings 50 and 51, and a connecting channel 53 likewise formed in the end wall. The openings 50 and 51 run close to the root circle of the toothing of the ring gear 2 within this root circle. Each of the two openings 50 and 51 is connected via a short, radially outwardly extending channel piece 54 and 55 to the circumferentially extending connecting channel 53, which is connected to the pressure kidney 20. The radial channel pieces, the openings 50, 51 and the connecting channel 53 are formed as grooves in the end wall of the gear chamber. For example, you can have a rectangular cross-section with rounded corners, the depth being approximately equal to the width of the groove shown. The connecting channel 53 is constantly covered by the ring part of the ring gear 2 which supports the teeth. Since shortly after leaving the point 16 of the tooth crest contact, the delivery cells slowly shrink, the end of the first opening 50 facing this point 16 can have a relatively large angular distance in the circumferential direction from this point, which is approximately equal to 2/3 of the angular dimension measured here Tooth division of the sprocket over this opening 50 is. In contrast, the end of the opening 51 located in the conveying direction is considerably further away from the front edge of the pressure kidney 20, namely slightly more than one tooth pitch, so that whenever a conveying cell loses contact with the opening 51, it immediately begins to get into the Open pressure kidney 20. The distance between the facing ends of the two openings 50 and 51 is so great that the two openings 50 and 51 are never connected by a conveyor cell; it can also be slightly larger if the openings are narrow.

When designing openings 50 and 51, the radial position of these openings must also be taken into account. Thus, in order to obtain the same opening and closing times, the extent of the openings 50, 51 in the circumferential direction must be smaller the more the openings are located away from the tooth root circle of the ring gear 2. In order to indicate this, the opening 50 is arranged somewhat further radially inwardly than the opening 51, but also extends somewhat less in the circumferential direction. In the exemplary embodiment, both openings 50 and 51 are relatively short, in many cases they are also made somewhat longer.

When the gerotor pump is operating at a low speed, the squeezed oil flow QL through the connecting channel 53 corresponds to the displacement volume of the delivery cells 17.1 to 17.3. With increasing speed, the flow apparent resistance for the flow through the connecting channel 53 now increases, since the open times for the openings 50 and 51 are becoming shorter and shorter. Accordingly, the pressure PI in the cells 17.1 to 17.3 increases with a simultaneous drop in the pinch oil flow QL through the connecting channel 53. However, these relationships only apply up to the speed at which no cavitation occurs in the suction kidney 11, that is to say in the delivery cells 13 . In the cavitation area at a higher speed, where the conveying characteristic curve (FIG. 3) has changed from a linearly increasing course to an approximately horizontal course, the pressures PI in the delivery cells decrease to near the atmospheric pressure. Since the intake pressure is kept constant over the speed, the QL curve now passes through the zero point and even becomes slightly negative. A small amount of oil flows from the pressure kidney 20 through the connecting channel 53 back into the delivery cells. At a very high speed, which almost never occurs in practice, the negative leakage oil flow QL from the pressure kidney 20 to the openings 50 and 51 would again approach the zero line due to the increase in the flow resistance. These relationships are shown in FIG. 6. FIG. 7 shows the corresponding suction pressure PS in the suction kidney 11 as a function of the pump speed, while FIG. 8 shows the intermediate pressure PI in the sealing web and the pressure difference PI-PH, PH is the pressure in the pressure kidney 20 as a function of the pump speed for such a pump.

The bypass formed by the openings 50 and 51 and the connecting channel 53 can also be provided in addition to the overflow channels 128 of the pump according to FIG. 4 provided with the check valves 21. This is even a preferred exemplary embodiment, since such a bypass additionally stabilizes the flow through the overflow channels 128 and counteracts a vibration of the valves 21.

FIG. 9 shows a cross-sectional view of an embodiment of an internal gear pump according to the invention. The pump has a housing 201 which encloses a gear chamber 206 with a toothed ring 202. A pinion 203, which has one tooth less than the toothed ring 202, meshes with the toothed ring 202. The pinion 203 forms, with the toothed ring 202, successive feed cells 210, 211, 212, 213, 214, 215 and 216 which seal against one another by the tooth engagement. An inlet channel 204 opens into an inlet mouth 207 designed as an inlet kidney, which is shown in broken lines. Furthermore, the inlet channel 204 in the position shown in FIG. 9 is connected via a housing bore 217 with housing shoulders 217a, 217b, 217c and 217d to the feed channels 222a, 222b and 222c, which run into the inlet openings 208a, 208b and 208c.

On the outlet side, the housing has an outlet channel 205, which is connected to the outlet kidney 209 arranged in the gear chamber 206, which is also shown in broken lines. Furthermore, the outlet kidney 209 is connected on its side facing away from the outlet opening 205 to an overflow channel 220 which opens into the housing bore 217 at the housing shoulder 217a on the opposite side of the inlet bore 204. A valve device is provided on the lower part of the housing 201. In this position of the valve device, a valve piston 221 is located in the housing bore 217, a head shoulder 224 of this valve piston 221 striking the housing with its front end face in the transfer channel 220 and the side bore of the housing bore 217 on the housing shoulder 217a against the liquid in the transfer channel 220 seals. At its rear end, the valve piston 221 is guided with its rear shoulder 229 in a spring chamber 225, in which spring 223 in the direction of the attachment point on the housing (in the left direction in FIG. 9) against the pressure in the transfer channel 220 or against the stop of the Head paragraph 224 on the housing 201. The spring chamber 225 is tightly closed at its right end with a locking screw, not shown. A Bore 226 in the valve piston 221 connects its surroundings with the spring chamber 225 filled with working fluid, as a result of which a damping effect occurs.

Starting from this FIG. 9, which designates all components, the mode of operation of the internal gear pump according to the invention will now be described with the aid of the other figures. Identical components are provided with corresponding reference symbols in all figures. For the sake of clarity, FIGS. 10 to 13 no longer refer to all, but only the relevant components.

In the state shown in FIG. 9, the pinion 203 is rotated in the direction indicated by the arrow n. Liquid is sucked in via the inlet channel 204 and, on the one hand, fed to the delivery cells 210 and 211 via the inlet kidney 207. On the other hand, working fluid is also supplied via the housing bore 217 in the space between the valve piston 221 and this housing bore to the feed channels 222a, 222b and 222c and via these to the inlet ports 208a, 208b and 208c, which supply the feed cells 212 and 213 with working fluid. In the state shown in Figure 9, the pump in the proportional range, i.e. the delivery rate increases linearly with an increase in speed n. Since the head shoulder 224 seals the housing bore 217 on the housing shoulder 217a against the liquid in the transfer channel 220, only the delivery cells 214, 215 and 216 are under pressure. The spring force F0 exerts a greater or equal pressure on the valve piston 221 as the pressure P₀ against the surface of the head heel 224 designated AK.

etwa konstant ist.In the following functional description it is assumed that a consumer is connected to the outlet channel 205, the hydraulic resistance of which $$\text{R =}\frac{\text{ΔP}}{\text{ΔQ}}$$

is about constant.

The regulation begins when the force exerted by the working fluid in the transfer channel 220 on the head heel 224 becomes greater than the spring force. In Figure 10, the pinion 203 rotates at the speed n1, which is already higher than the limit speed in the proportional range of the pump. The pressure of the working fluid in the pressure range would increase linearly to a pressure P _{1 '} , so that the valve piston 221 is moved to the right. As a result, the suction angle α _s is reduced from α _{s max} (see FIG. 9) to α _s1 (see FIG. 10). The pressure P _{1 '} , which would be reached linearly, can not hold, but drops to P₁. This means that the flow rate also drops linearly. There is a new flow rate and a new pressure P₁ at the increased speed n ₁ , which is lower than P _{1 '} , but higher than P₀. The setting of a pressure P₁, which is higher than the pressure P₀, is also structurally due to the design of the valve device and the pump. If this pressure were not higher than P₀, the valve piston 221 would be pushed back into the original position (FIG. 9) by the spring 223, and the process would start again because the speed is increased compared to the initial position. Had the pressure P 1 remained in the pressure range at the value P 1 ', the throttling action of the piston 221 moving to the right would have been ineffective due to the head shoulder 224 penetrating into the feed channel 222 a on the filling of the feed cell 212. Thus, the pressure P₁ must be between P₀ and P₁ '.

A summary of FIGS. 10 and 11 shows what happens when the speed is increased further, here to the speed n 2 in FIG. 11. The process described above for increasing the speed continues, so that the valve piston 221 keeps increasing to the right due to the pressure increase 11, until a state is reached, for example, as shown in FIG. 11, where the valve piston 221 seals the housing bore 217 with its head shoulder 224 on the housing shoulder 217c, so that the feed cell designated here 212 does not supply suctioned-in working fluid via the inlet channel 204 but via the transfer channel 220 and the channels 222a and 208a with pressurized working fluid. The working fluid in the feed cell 212 is with the downstream feed cells at the increased pressure P₂, so that no cavity is created in it and no negative pressure can develop despite the increase in space. On the contrary, this supply cell 212 generates a positive torque on the pinion 203 by the pressurization with the pressure P₂, because its space expands under high pressure and works like a hydraulic motor. This internal differential control thus works with high efficiency. The working fluid under pressure P₂ is not decompressed to atmospheric pressure, but instead returns its potential energy as mechanical power to the pump drive shaft with a certain loss of flow through the channels. The suction angle in this position is labeled α _S2 .

In the state shown in Figure 12, the speed n₃ is now increased so far that the valve piston 221 has moved so far to the right that it seals the entire housing bore 217 with its head shoulder 224 against the working fluid in the inlet channel 204 on the housing shoulder 217d. The delivery chamber designated 212 and all of the downstream delivery chambers from it are now supplied with pressurized working fluid either via the outlet kidney 209 or via the transfer channel 220 and the supply and inlet channels 222a, 222b, 208a and 208b crossing them. The spring 223 is pressed onto the block. Half of the feed cells used in the initial stage for suction are separated from the inlet channel 204 and at the same time connected to the high pressure P₃, so that they are effective as a hydraulic motor, as described above. Above all, the pump works practically without cavitation in the entire regulated area, so that there is no noise. In the speed range from n₀ to n₃ in the inlet channel 204, no aperture or other throttle is necessary because of the internal control just described.

If, as in FIG. 12, the valve piston 221 is pressed to the right up to the spring block, no further internal regulation can take place. With further increases in speed, the delivery rate will continue to increase in proportion to the speed with reduced steepness, until cavitation occurs in the remaining remaining suction tooth chambers in the region of the short suction kidney 207.

The pump described above is mainly suitable for supplying automatic transmissions with a pressure level of up to 25 bar or higher. The stiffness of the spring 223 determines the steepness of the conveyor line in the regulated area and must be adapted to the hydraulic resistance of the consumer.

FIG. 13 shows a further embodiment of the internal gear pump according to the invention, in which two further aspects of the present invention emerge. A first aspect relates to the design of the pump with a pinion 203, which has two teeth less than the toothed ring 202.

At the point at which the teeth of the pinion 203 disengage from the toothed ring 202, a crescent-shaped filling piece 227 is provided here. The teeth 228 of the toothed ring 202 are designed to be sufficiently pointed to sufficiently seal the feed cells for the tooth engagement in the suction area.

The operation of the internal gear pump shown in FIG. 13 and the function of the valve device correspond to those described in FIGS. 9 to 12.

Another aspect of the invention, which becomes clear from FIG. 13, relates to the safety valve effect of the valve device. This works as a bypass valve when, at maximum pressure in the pressure area, the head attachment 224 has exceeded the last feed channel 222c to such an extent that a short circuit from the pressure area enters the inlet channel 204 under decompression. The spring 223 is only allowed to block when an outflow cross section sufficient for this purpose has been reached at this point. For the function of the valve piston 221 as a safety valve, the head extension 224 must be longer than the width of the recess 230. In FIG. 13, the head extension 224 is designed accordingly. If the head approach is too short, the piston loses its guidance.

As is also shown in FIG. 13, the head shoulder 224 of the valve piston 221 here consists of a heel base 224a and a heel tab 224b which adjoins this and has the same outer diameter. The guidance and the sealing function of the valve piston 221 in the housing bore 217 on the housing shoulders take place on the outer surfaces of the heel base 224a and the heel tab 224b. Although the heel base 224a is itself narrow, in particular narrower than the width of the feed channels 222, good routing and sealing can be ensured by the milled out shoulder tab 224b.

Claims (29)

Valve control of an internal combustion engine a) with a hydraulically actuated adjusting means for adjusting a valve control means depending on the engine speed and b) with a pump (100) driven by the motor for supplying the actuating means with working fluid, characterized in that c) the pump (100) is designed as a suction-controlled gerotor pump with a sealing web extending over several delivery cells (17.1 - 17.3) and has a speed-dependent delivery characteristic curve which is adapted to the working fluid requirement of the actuating means. Valve control according to claim 1, characterized in that the valve control means are camshafts, the phase position of which can be changed in order to control overlap times of intake and exhaust valves. Valve control according to claim 1 or 2, characterized in that the pump (100) supplies an adjusting means for changing a valve stroke with working fluid. Valve control according to one of the preceding claims, characterized in that the pump (100) supplies an adjusting means with working fluid with which a cylinder of the engine can be switched on and off. Valve control according to one of the preceding claims, characterized in that the pump (100) supplies the motor with lubricating oil and the lubricating oil also serves as working oil for the hydraulic actuating means. Valve control according to one of the preceding claims, characterized in that a throttling (14; 43) of the pump (100) on the intake side can be changed in order to be able to adapt the delivery characteristic of the pump to the requirements of the valve control. Valve control according to the preceding claim, characterized in that the throttling (14, 43) is designed in stages, in particular in two stages. Valve control according to one of the preceding claims, characterized by a pump - a housing (1), - An internally toothed ring gear (2) arranged rotatably in a gear chamber of the housing (1), - A tooth less than the ring gear (2) with the ring gear (2) meshing, arranged in this pinion (4), the Teeth together with the teeth of the ring gear (2) form increasing (13) and decreasing (17) delivery cells, which follow one another and are sealed off from each other and each with the adjacent delivery cells by in the ring gear (2) and / or the pinion (4 ) provided overflow channels (128) are connected, Non-return valves (21) in the overflow channels (128), which counteract a flow of the working fluid against the conveying direction, - In the housing (1) arranged inlet and outlet channels (12, 19) for the supply and discharge of the working oil, which open into the gear chamber on both sides of the point (7) deepest meshing, one of the point (7) deepest meshing the distal end of a mouth (20) of the outlet channel (19) is so close to the point (7) of the deepest tooth engagement that between it and the circumferential point at which the delivery cells begin to decrease, there are always a number of decreasing delivery cells (17.1 - 17.3). Valve control according to the preceding claim, characterized in that - The mouth (20) of the outlet channel (19) is preceded by at least one further outlet (35) connected to the outlet channel (19) in the circumferential direction of the pump, which is connected via a line (36) to the outlet channel (19), - The flow through this line (36) by means of a throttle element (37) is controllable, in particular lockable, and that - A control device for the intake-side throttling (14, 43) and the throttle element (37) is provided. Valve control according to the preceding claim, characterized in that the control device restricts the intake (14, 43) and the throttle element (37) controls the working fluid requirement of the adjusting means. Valve control according to one of claims 8 to 10, characterized in that - In the area of the shrinking delivery cells (17.1 - 17.3) in a wall of the gear chamber in the circumferential direction at a distance from the mouth (20) of the outlet channel (19) at least one opening alternately swept by delivery cells (17.1 - 17.3) and teeth delimiting these (50 , 51) lies, - The opening (50, 51) via a connecting channel (53) with the outlet channel (19) is connected, and that - The opening (50, 51) is completely or at least largely covered by a tooth each time it overflows. Valve control according to Claim 11, characterized in that a plurality of openings (50, 51) are connected to the outlet duct (19) via the common connecting duct (53). Valve control according to claim 11 or 12, characterized in that the connecting channel (53) has a channel piece (54, 55) branching off radially from the opening (50, 51). Valve control according to one of Claims 11 to 13, characterized in that, if there are a plurality of openings (50, 51) arranged one behind the other in the circumferential direction, these are at a distance of approximately half a tooth pitch from one another and the extent of the opening (50, 51) in the circumferential direction is approximately equal to that Thickness of the teeth sweeping them over the radial height of the opening (50, 51). Valve control according to one of claims 11 to 14, characterized in that the openings (50, 51) radially extend from approximately 1/5 to approximately 1/3 the height of the teeth sweeping over them. Internal gear pump with a) a housing (201) with a gear chamber (206), b) a toothed ring (202) in the housing (201), c) a pinion (203) which is arranged in the toothed ring (202) and meshes with it and which has at least one tooth less than the toothed ring (202) and together with this successive conveyor cells (210, 211, 212, 213) which seal against one another by the tooth engagement, 214, 215, 216) for the working fluid, and d) at least one inlet channel (204) and at least one outlet channel (205) for the working fluid in the housing (201), e) wherein the working fluid is fed from the inlet channel via at least one inlet opening (207, 208a, 208b, 208c) into the suction area of the gear chamber (206) and via at least one outlet opening (209) from the pressure area of the gear chamber (206) into the outlet channel (205) is removed, marked by f) a device (220, 221, 222) which, as the pressure in the pressure range rises, feeds a regulated amount of the working fluid from the outlet mouth (209) into at least one inlet mouth (208a, 208b, 208c) while simultaneously supplying working fluid interrupts the inlet channel (204) into this inlet mouth (208a, 208b, 208c). Internal gear pump according to claim 16, characterized in that the device (220, 221, 222) as the pressure in the pressure area increases, the inlet openings (208a, 208b, 208c) connects to it. Internal gear pump according to claim 16 or 17, characterized in that the device (220, 221, 222) has a transfer channel (220) connected to the outlet mouth (209), which via a valve device (221, 222, 223) into at least one supply channel ( 222a, 222b, 222c) which in turn communicates with an inlet mouth (208a, 208b, 208c). Internal gear pump according to claim 18, characterized in that the valve device (221, 223, 224) has a valve piston (221) which is supported against the pressure of the working fluid in the transfer channel (220) by means of a spring (223) supported in the housing (201) and blocks or opens the access of the working fluid into the supply channels (222a, 222b, 222c) by means of a head heel (224). Internal gear pump according to claim 19, characterized in that the valve piston (221) in the depressurized state of the transfer duct (220), or up to a predetermined pressure therein, against the force of the spring (223) by a stop on the housing (201) in is held in a position where no working fluid flows from the transfer channel (220) into a supply channel (222). Internal gear pump according to one of Claims 19 or 20, characterized in that the valve piston (221) is stopped in its position against the direction of the spring force in the position where working fluid flows from the transfer channel (220) into all supply channels (222), that the spring (223) is blocked. Internal gear pump according to one of claims 16 to 21, characterized in that the inlet mouth (207) for the delivery cells (210, 211) not to be connected to the transfer channel (220) is in its Size is limited to about the area in which these conveyor cells extend. Internal gear pump according to one of claims 16 to 22, characterized in that the outlet opening (209) extends approximately over the entire area of the conveyor cells (214, 215, 216) which are downstream of the conveyor cells (212, 213) in the conveying direction can be connected to the transfer channel (220). Internal gear pump according to one of claims 19 to 23, characterized in that the end of the valve piston (221) facing away from the head shoulder (224) forms together with the housing (201) a spring chamber (225) which is filled with working fluid to dampen the piston movement and is in fluid communication with the working fluid in the inlet channel (204) via a bore (226). Internal gear pump according to one of claims 16 to 24, characterized in that the valve device (221, 223, 224) simultaneously acts as a safety valve in the form of a bypass valve when the head shoulder (224) exceeds the last supply channel (222c) so far at maximum pressure in the pressure range has that a short-circuit flow of the working fluid occurs from the pressure area into the inlet channel (204) under decompression. Internal gear pump according to one of claims 16 to 25, characterized in that the pinion (203) has two teeth less than the toothed ring (202) and a crescent-shaped, housing-fixed filler is provided at the point of disengagement of the teeth. Internal gear pump according to claim 26, characterized in that the teeth of the toothed ring are designed to be sufficiently pointed so that the feed cells (210, 211, 212) in the suction region are meshed with the teeth are sealed against each other. Internal gear pump according to one of Claims 19 to 27, characterized in that the head shoulder (224) of the valve piston (221) consists of a shoulder base (224a) and a shoulder lug (224b) which adjoins this and has the same outer diameter, the guidance and the sealing function of the valve piston (221) in the housing bore (217) on the housing shoulders (217a, 217b, 217c, 217d) on the outer surfaces of the heel base (224a) and the heel tab (224b). Internal gear pump according to one of claims 16 to 28,
marked by
the use as a suction-controlled gerotor pump of a valve control according to one of claims 1 to 7.

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