EP0678166A1

EP0678166A1 - Control device for a variable intake volume pump

Info

Publication number: EP0678166A1
Application number: EP94930902A
Authority: EP
Inventors: Wolfgang Schneider
Original assignee: EIDGENOESSISCHE TECHNISCHE HOCHSCHULE LABORATORIUM fur VERBRENNUNGSMOTOREN und VERBRENNUNGSTECHNIK; Eidgenossische Technische Hochschule Laboratorium fur Verbrennungsmotoren und Verbrennungstechnik; Eidgenoessische Technische Hochschule Zurich ETHZ
Current assignee: CRT Common Rail Technologies AG
Priority date: 1993-11-08
Filing date: 1994-11-07
Publication date: 1995-10-25
Anticipated expiration: 2014-11-07
Also published as: ES2120076T3; EP0678166B1; CN1116441A; JPH08505680A; ATE169720T1; JP3747061B2; CN1082143C; US5701873A; DE59406680D1; CA2151518A1; WO1995013474A1

Abstract

PCT No. PCT/CH94/00215 Sec. 371 Date Jul. 10, 1995 Sec. 102(e) Date Jul. 10, 1995 PCT Filed Nov. 7, 1994 PCT Pub. No. WO95/13474 PCT Pub. Date May 18, 1995A control device for a filling-ratio adjusting pump with at least one displacement space works on the suction-throttle principle with a positive variation in volume of the displacement space or displacement spaces and is intended inter alia particularly for common-rail diesel injection systems. It allows an exact, precise and highly dynamic control of the filling-ratio adjusting pump at low outlay, without the system being impaired by undesirable cavitation. Located on the suction side of the pump is at least one throttling 2/2-way valve (21, 21a, 21b; 134; 51, 52, 53, 54; 81; 103) actuated by pressure difference. Either such a 2/2-way valve can be used for a group of displacement spaces or for the entire pump or a respective valve of this type can be inserted in front of each individual displacement space. The pressure-difference control of the or each 2/2-way valve takes place via an adjusting device (27; 150) which is arranged on the inflow side of the 2/2-way valve and which is designed either as a throttling valve or as a flow-regulating valve.

Description

Control device for a filling level variable pump

description

The invention relates to a control device for a filling _. Grad-variable pump with at least one displacement chamber, which works according to the suction throttle principle with an inevitable change in volume of the displacement chamber or displacement chambers, which pumps the liquid to be delivered from a liquid reservoir with a free surface which is acted upon by a gas pressure, usually atmospheric pressure , through a line or possibly via a hydraulic system but without gas supply.

Filling level variable pumps are hydrostatic pumps with displacement effect through reciprocating pistons (e.g. radial piston pump, axial piston pump, in-line pump) or rotary or swinging piston pumps (e.g. vane pump, vane pump, roller cell pump). The invention relates only to such filling degree variable pumps that operate on the principle of suction throttling with an inevitable displacement movement. In these, the displacement space is partially filled by controlled cavity formation in the pressure fluid. Inevitably moving displacers can be considered both pistons with an oscillating movement and rotary displacers (vane cell pump, blocking vane pump, etc.).

To increase the energy efficiency in hydrostatic systems, there has long been a desire for increased use of variable pumps. The designs of such variable pumps available today, which are mostly implemented according to the principle of stroke adjustment, are still too expensive for many applications or have an inefficient efficiency with partial delivery, ie with a low degree of filling. At the same time, due to the undiminished increasing cost-benefit ratio of electronics, the trend of linking electronics and fluid technology continues, so that there is an increasing call for direct but nevertheless inexpensive electrical control of variable pumps.

For the integration into control systems (in the form of actuators), future variable pumps must have defined flow characteristics and reproduce them precisely, with little hysteresis and with sufficient speed (e.g. without great dead time). As is well known, such properties are, in part, indispensable for control elements in control loops, in some cases at least of considerable advantage.

Furthermore, a high level of uniformity of conveyance between the individual displacers is important, on the one hand because of the noise excitation and because of any consumers who are dependent on uniformity, and on the other hand so that no additional disturbances of different frequency are carried into the high-pressure system, which could irritate a regulator.

Hydrostatic filling level adjustment pumps of this type can be used in many fields of application in vehicle, industrial, flight and water hydraulics as well as especially for general automotive hydraulics and the so-called common rail diesel injection systems. By using the phase gating principle in such filling degree variable pumps (see bibliography at the end of the description), very high efficiencies can also be achieved with partial conveying, and in particular even with low-viscosity media, very high pressures and the lowest speed. Unlike in the case of the stroke variable displacement pump, the duration of the pressurization of the displacer body and that with it is reduced in the case of the phase-controlled pump with decreasing delivery rate per work cycle associated loss work (such as piston gap leakage). This property of being insensitive to leakage leads, among other reasons (see FIGS. 2 and 4 of the bibliography at the end of the description), to the particular suitability of such pumps for common-rail diesel injection technology.

Such a reason is also the low expenditure of energy or force for the adjustment, since this is often done by adjusting a throttle in the low-pressure part (US Pat. No. 4,907,949). This also enables manual adjustment, among other things.

In principle, the low expenditure of force also enables a very high adjustment dynamics, so that the necessary adjustments are not only calculated quickly by electronic means, but that the adjustments can be realized by using fast components for the electrical direct drive. Because of the low forces, the size and manufacturing costs of the electric drives are also low. In general, low forces allow the control of hydraulic-mechanical systems with largely no feedback from the manipulated variable from the measurement signal.

The common-rail diesel injection system is an example of the high adjustment dynamics required; The manifold (= common rail) and the other high-pressure-carrying volumes must be able to be inflated to a considerably higher pressure very quickly (in the order of 0.2 seconds when used in an automobile) when the engine electronics signal. For this purpose, the delivery rate of the pump must be able to be adjusted an order of magnitude faster - a pump cycle can be reached as little as possible. For this, you can again refer to 4) of the bibliography. Such a pump also has to be constant Pressure on the order of about two injections can provide other flow rates.

Other previous solutions, e.g. for pumps with a larger number of displacement spaces, are with individually controlled inlet valves, too expensive. It is a great advantage to get by with just one adjustment element for a large number of cylinders.

PCT / EP89 / 01057 discloses an example of a generic control device for filling level variable pumps, which manages with only one converter element for a large number of displacement spaces and has an inlet-side slot control, as is sufficient for many applications. A special flow guide in the eccentric housing is intended to achieve a uniform filling of all displacement spaces and thus a high level of delivery even with partial delivery which is sufficiently small for many applications. However, the dynamics are not sufficient for various applications, since all cylinders are filled from the central eccentric housing, and this must first be filled by the throttle element in the case of a direct transition to full delivery and must be emptied in the reverse process before the filling and the delivery flow returns to a steady state. However, surface and gravity effects can still produce sporadic local bubble accumulations with subsequent group-wise loosening of walls, for example, which can lead to statistical delivery rate spreads and hysteresis effects in pump operation. For example, conditions can arise in which one displacer receives more liquid and the other displacer receives more void fraction during the filling, and as a result the delivery is also uneven. The dynamics of a pump similar to that in the aforementioned patent was measured by Fassbender (see 6 in the bibliography). In this case, the flow rate lags by a dead time of approx. 7 work cycles compared to the movement of the actuator. So a fast actuator alone is not enough. In the common rail diesel injection technology already used as an example above, the pressure in the distributor pipe would already increase inadmissibly during this period due to its small volume and would be difficult to regulate.

Another example of a generic control device for a pump with inlet valves implemented as check valves is known from CH 674 243 = EP-A-299 337. This published state of the art does not give any information on the prints used. When examined pumps of this type, however, it was found that they suffer from cavitation in the suction throttling, so that considerable gas volumes arise which significantly impair the desired precise, precise and simple control.

Due to the formation of cavities in suction throttling with an inevitable displacement movement, the throttle adjustment element was placed close to the displacement space in order to achieve the desired high dynamics. Consequently, at least in the case of the radial piston type, one converter is required per displacement space, or a complex mechanical linkage. The single-cylinder design was preferred, and a multiple cam or a gearbox was optionally proposed to achieve a higher volume flow and higher pump frequency. Apart from the associated structural and possibly load-related restrictions, such a solution with n-fold cams or n-fold drive ratio results An n-cylindrical pump has a high periodicity, i.e. a high similarity of the individual delivery processes, but also a greater degree of interruption, significantly (n times) higher torque peaks in the drive, larger noise emissions due to (n times) steeper pressure increases in the cylinders and the Danger that the re-entry process of gas molecules from the cavities back into the liquid can no longer keep up with the speed of the pressure increases (see 1 in the literature listing) and cavitation damage can then occur under this condition.

The company Cooper Bessemer Corporation, Mt. Vernon, Ohio, USA built a two-cylinder piston pump of the generic type for common rail diesel injection systems for many years. This pump had two cylinders, the variable throttle element being arranged between the two cylinders, so that the cavities fillable harmful spaces were minimal. Here too, expanding to more than two cylinders is difficult and time-consuming. The position of the adjusting throttle element between the two cylinders limits the freedom in the displacement arrangement (radial, axial, in series). This pump was again equipped with inlet slots, whereby a satisfactory tightness of the displacement spaces was apparently achieved by long stroke (i.e. correspondingly large sealing lengths and smaller gap lengths), but this required a correct crankshaft with cup tappets that absorb transverse forces and significantly increased the construction volume.

In summary, it can be said (firstly) that the disadvantage still exists today that for optimal dynamics, exactness of the delivery characteristic and freedom from hysteresis, properties as are particularly desirable when using the pump in controls, each displacement chamber has its own control element with drive must be equipped or the control elements have to be connected to a central drive element by means of complex mechanics, with the corresponding problems of quantity adjustment. This conflict of goals between simplicity (as few or only one control element as possible) and high dynamics, exactness of the conveying characteristic and freedom from hysteresis are all the more apparent when the individual displacement spaces, for example in the case of radial or row arrangements, are far apart or when the Displacement space is large. In the case of closely spaced displacement spaces such as axial piston pumps, a central arrangement of an adjusting element would in principle be conceivable, but the installation space is often too narrow or provided for other components.

The reason for these various restrictions when using the filling control by suction throttling with inevitable displacement movement lies in the void formation that was previously necessary for adjusting the delivery rate, which due to the turbulence that is always present and is sometimes desired for the purpose of viscosity independence, usually already in the throttling direction uses the switching device, and not only in the displacement rooms.

The object of the invention is therefore to create an inexpensive control device according to the preamble, which at least considerably reduces the effect of this obstacle to premature cavity formation with little effort and is thus generally applicable to various pump types of the displacer type of different sizes Degrees of freedom in the implementation of this actually very interesting and promising flow control can help. The creation of degrees of freedom is understood to mean that - from the point of view of manufacturing costs, the general applicability mentioned for various pump types, sizes and Design of the entire pump - it should be possible to combine control elements and, for example, actuate them directly using an electromechanical converter, and to be able to arrange the adjustable elements anywhere in the pump without significant deterioration in properties, or even to place them at a certain distance from the pump can, which gives a remote control possibility.

The formation of cavities in liquids in the case of stationary flows and the associated cavitation damage have been extensively investigated in the past. However, the unsteady and virtually non-flowing case of cavity formation in pump cylinders has only been investigated to a lesser extent. Apparently, however, damage caused by cavity disintegration is not to be expected with materials customary in pump construction. One of several reasons may be that the time is too short to release really large amounts of gas or steam. Schweitzer (see please 5 in the bibliography) examined the escape of dissolved gases from liquids and found diffusion time constants that are clearly above typical working cycle periods of hydrostatic pumps. Fassbender (6) in the bibliography) measured gas outlet pressures, and these are actually very low for many relevant liquids.

The present invention uses these physical phenomena and the further, more well-known fact that liquid which has been given time to saturate in a gas atmosphere of a pressure p1 (for example when resting in a tank ventilated to the atmosphere), when it falls below this Pressure - especially if turbulence is added when flowing through or around an obstacle - has a strong tendency to get rid of the excess gas. This may be little in terms of volume, but nevertheless a large part of a line or volume fill a volume len, which makes the above-mentioned filling or emptying processes necessary in terms of dynamics until a new steady state arises.

According to the invention, a control device according to claim 1 or according to claim 15 is provided to solve the above-mentioned problems. The main characteristic of the invention is an upstream connection of passive, according to the rules of the claims, throttling valves in front of the individual displacement spaces, in front of groups of displacers or the entire pump, which ensures that the pressure behind a throttle actuator up to these valves the pressure pl des Liquid _reservoirs and preferably pl plus does not fall below an amount Δ _pTemp , which is explained later, at least substantially, and thus a noteworthy and disturbing cavity formation is limited to the comparatively small volume behind these valves up to the displacement spaces. This procedure is unusual since throttling under pressure loss in pumps is otherwise avoided as much as possible, for example by not or only slightly pretensioning inlet valves in order to maintain a certain self-priming capacity of the pump or the risk of cavitation in the suction line , for example at kinks. For this purpose, the pressure p3 upstream of the throttling valves generally has to be raised slightly by a pressure source of a known type, for which purpose a height difference from the liquid reservoir to the valve inlet would also be possible. For this reason, most hydraulic systems, but especially hydraulic and fuel supply systems in vehicles, work with pumps which generate low to very low admission pressures, so that practically no significant restriction for the application or introduction of the invention arises from this condition.

An important finding on which the invention is based lies in the fact that at atmospheric pressure a liter of fuel or hydraulic fluid is able to absorb about 10% volume of air in dissolved form and does so in a fuel tank of a vehicle. At atmospheric pressure, a gas volume of around 100 cc is therefore contained in one liter of fuel. When the pressure is reduced, this dissolved air comes out of the solution in gas form and expands in volume in accordance with the prevailing negative pressure, for example at 0.1 bar, the gas volume increases tenfold to 1000 cc. Such gas volumes can very quickly fill the existing volume behind the valves up to the displacement spaces and thereby severely impair the delivery and control of a fuel pump. The same consideration applies to other liquids. Due to the inventive limitation of Δ _p6min to not less than 0.9 bar and preferably in the range between 1.0 and 1.5 bar, the formation of gas volumes is kept to a minimum or completely prevented, so that the delivery and control of the pump system does not suffer.

The rules in claims 5 and 6 take into account the specific properties of the liquids and gases. The formula according to claims 5 and 6 enables the determination of the minimum opening pressure difference Δ _pömin at which the or each pressure difference-actuated throttling 2/2-way valve opens both for pumps with inlet slots and for pumps with automatic, spring-loaded, displacement-controlled inlet valves. If gas outlet pressures pGasaus and steam outlet pressures pDampfaus are not known, the formula is on the safe side with 0 bar for these pressures.

The so-called solubility coefficient describes liquid-specific and gas-specific the solution behavior according to the equation by Henry: it = k

saturation concentration of the dissolved gas or gas mixture in the liquid

P Pressure at saturation equilibrium (pl) k = k (T) Solubility coefficient of the gas or gas mixture in the liquid.

In many plants, especially e.g. In vehicle use, for example on the way from a still cold tank to an already hot engine, a liquid to be conveyed can experience a rapid change in temperature.

If the solubility coefficient becomes smaller in the direction of the temperature change, this can lead to a sudden oversaturated state of the liquid, which can lead to disruptive gas release even before the throttling, spring-loaded valve.

To reliably prevent this, i.e. To at least maintain the saturation concentration, the maximum temperature-related drop in the solubility coefficient k that occurs during operation can be prevented by increasing the minimum opening difference by a Δptemp.

If it .. = it, then with Henry k (T _x ) p _x = k (T ₁ ) p ₁ p _χ / _Pl = k (T _χ ) / k (T _x )

or Δptemp = p _x - p ₁ = (p _x / P _! - 1) p _x = (k (T ^ / k (T _χ ) -l) p ₁ if k (T _χ ) <k (T ₁ ) , in which T _χ and T _{x determine} the maximum temperature difference of the liquid between the liquid reservoir and the throttling, spring-loaded valve that occurs during operation at intervals of a few hours.

The main advantage of the control device chosen according to the invention is the desired fast, reproducible, low hysteresis and low dead time reaction of the delivery rate to adjustments of the actuators. This exact, calculable assignment of actuator position and pump flow is in turn a prerequisite for the integration of this pump in control loops of hydraulic systems, especially in those with high demands on the control dynamics, such as those that exist for common rail diesel injection systems, among other things. In the case of theoretically infinitely fast actuation of the actuator, the assigned, full response to the delivery rate takes place with the first subsequent, complete suction process (in principle, it couldn't be any faster). In this way, the pump delivery flow can already be changed at the same time in a hydraulic system if an expected sudden change in consumption is known.

The great lack of cavities up to the throttling valve close to the displacement space (and in the special case, if this is designed as an inlet valve, to the limit of the displacement space) allows the use of various adjusting devices for filling control and many advantageous specializations, since the most varied pumps have the Displacement type degrees of freedom can be obtained.

Particularly preferred embodiments of the control device according to the invention for a filling level adjustment pump can be found in the further subclaims.

The invention is described below with reference to exemplary embodiments. play explained in more detail with reference to the drawing, in which:

1 shows an embodiment of a control device according to the invention for a pump with automatic inlet valves,

Fig. 2 shows another embodiment of an inventive

Control device for a pump with inlet slots controlled by the displacer,

Fig. 3 shows a special embodiment of an inventive

Control device for a pump, the inlet valves being designed with a special spring characteristic and a damper and the adjusting throttles being combined in a continuous way valve,

4 shows a cross section of an executed pump with a control device according to the invention which is installed in the pump,

Fig. 5 is a partially longitudinal sectional schematic

4 along the line V-V in Fig. 4,

6A and 6B drawings to explain the mode of operation of the inlet valves of the pump of FIGS. 4 and 5, FIG. 6A representing the opening process and FIG. 6B the closing process,

7 is a drawing to explain the design of the characteristic of the throttling valves,

8 is a graphical representation of the operating cycle of a pump according to FIG. 3 for full delivery, 9 and 10 representations corresponding to FIG. 8, but for half funding or zero funding,

11 flow characteristics of a pump of FIG. 3,

12 flow characteristics similar to FIG. 11, but for a slot-controlled pump,

13 shows an embodiment of a control device according to the invention with a switching valve as an adjusting device,

14 shows an embodiment of a control device according to the invention for a filling degree variable pump, in which the adjusting device is formed by a variable displacement machine,

15 is a similar embodiment to that of FIG. 14,

Fig. 16 shows a further variant of an inventive

Control device for a filling level variable pump, in which an adjustable pressure relief valve serves as an adjusting device

17 shows a preferred variant of a control device according to the invention for a filling degree distribution pump, in which the adjustment device is equipped with an auxiliary medium, i.e. does not work with the liquid to be pumped, and

18 shows a schematic view of a further degree of filling variable pump according to the invention.

1 shows a first possible embodiment of a control Direction for a pump with automatically operating inlet valves.

The pump according to the schematic representation of FIG. 1 has three individual displacement pistons 9, only one of which can be seen in FIG. 1. The three displacers ben angetrie¬ of a ^"rotation shaft 12 via respective eccentric 11, each cam 11 is disposed in a lifting member 10, which is located at the lower end of the associated piston. 9

In this case, the rotary movement A of the eccentric 11 initiates an oscillating movement B, the piston 9, as a displacer in the displacement space 15, moving back and forth between the two dead center positions C (bottom dead center) and D (top dead center) and triggering the periodic suction movement. The piston 10 does not lift off the eccentric 11 in any phase of its movement (inevitable displacement movement). An inlet valve 28 and an outlet valve 17 are provided for each displacement chamber in a manner known per se, whereby both the inlet valve 28 and the outlet valve 17 can be biased into the respectively closed positions by respective springs (for example 29 for the inlet valve 28). This means that the valve 28 is designed as an inlet check valve. Due to the movement of the displacer 9, due to the rotary movement of the eccentric 11, the inlet check valve is turned on in a known manner via the pressure difference p4-p5 which arises and the suction process is triggered. During the upward stroke of the displacer 9, the previously collected amount of liquid is displaced from the displacer space 15 by the outlet valve 17, that is, it rises from its seat against the action of the biasing spring and the liquid which is now under high pressure is supplied via line 18 with corresponding Amounts of liquid Conveyed via lines 18a and 18b into a common line 19, where there is a pressure p6 and which represents, for example, the so-called "common rail" (the distribution pipe) of a "common rail" injection system.

As is customary in such multi-piston arrangements, the individual pistons or displacers 9 are moved out of phase in order to achieve an equalization of the outlet pressure p6 in the common line and to ensure that the pump is operated with as little vibration as possible. That is, in the case of three displacers, as shown in the example according to FIG. 1, the individual displacement pistons each carry out their stroke movement out of phase with the neighboring displacer by 120 °.

The flow rate through each displacer is determined by a respective, upstream, throttling, spring-loaded 2/2-way valve 21 and by an adjusting device 27, which in this example is designed as an adjusting throttle 30.

The adjusting device 27, like the adjusting device 27a and 27b of the same design, is fed by a common line 32, which provides the liquid to be pumped, here diesel oil, with a pressure p2. The diesel fuel 2 comes from a liquid reservoir 1, where it is in contact with a gas 3 at a pressure p _x here air at atmospheric pressure pl at a contact surface 4. The liquid can become saturated with gas. The liquid initially flows through a system 7, in which preferably no further gas is to be added to the liquid. Since the pressure is to be increased from p1 to p2, a pressure-increasing device, ie a pressure source 8, is integrated in the system 7 in this example.

The diesel fluid then flows through line 32 the three adjustment throttles 30, 30a, 30b and the throttling 2/2-way valves 21, 21a and 21b assigned to them and operated by pressure difference. Due to the continuity equation for incompressible media (which can only be assumed on the basis of the freedom from voids), the flow rate through each adjustment throttle and the 2/2-way valve 21, 21a or 21b ^{"assigned to it is the} same. This results in an equilibrium state the pressure p3 on the active surface 24 of the 2/2-way valve 21 on one side and a reservoir-like pressure pl2 near pl on the active surface 23 on the other side of the 2/2-way valve and from the opening-path-dependent force of the spring 22. The adjusting throttles 30, 30a, 30b can theoretically be individually adjusted or matched to one another.

1 discloses another important advantage of the invention. With the valve active surfaces 24, 24a, 24b and the associated throttles 30, 30a, 30b, the system has an inherent damping effect which increases with higher throttling, which ensures compliance and reproducibility of the delivery characteristics (see FIGS. 10 and 11 ) important is. the damping works by already producing a slight overshoot of the throttling 2/2-way valves 21, 21a, 21b in the opening phase that suddenly begins, which produces the product of the surface 24, 24a, 24b and the stroke difference in the connection 31, 31a, 31b Volume increase causes a decrease in pressure p ₃ by a considerable Δp ₃ which counteracts overshoot - because of the lack of voids according to the invention!

The smaller the throttle is set, the longer it takes for medium to flow and the more sustainable the damping effect.

In this exemplary embodiment, the throttling 2/2-way valves 21, 21a and 21b actuated by pressure difference are on its active surface 23 connected to the return 6, whereby the reservoir-like pressure pl2 prevails near pl on the active surface 23.

This arrangement has the advantage that the spring 22 - depending on the size of the surface 23 - can be chosen to be very weak and serves less for preloading than for the regulating resetting of the valve (21) against the opening pressure p3 on the other active surface 24, since with the pressure pl2 on the active surface 23 there is already a considerable part of the necessary pretension and maybe even more.

FIG. 2 shows a similar control device to FIG. 1, with the difference that the pump has inlet slots 35 and only one central adjusting device 27 is provided, which has an adjusting throttle 30. Pumps with inlet slots can generally be manufactured more cost-effectively than those with inlet valves, while their main application is less for the highest pressures and low-viscosity pressure media.

In this exemplary embodiment, the low cost target is accommodated by the central adjustment device 27, which in principle permits simple manual adjustment or electrical adjustment. The individual adjusting throttle 30 in the adjusting device 27 can also be inexpensively represented in a manner known per se. In the suction phases, the pressure difference p2-p3 above the adjustment throttle is kept approximately constant, regardless of the flow rate, by means of a pressure difference valve 40 connected in parallel, whereby the effect of a flow control valve results in the combination of the adjustment throttle 30 and the pressure difference valve 40. The simplicity of using the same adjusting throttle 30 for all displacement elements 16, 16a, 16b gives further advantages in this constellation with the inlet line slit control of the pump.

A first advantage is that the control cross section of the throttle 30 for a certain speed and certain relative displacement space filling is significantly "larger" than the individual throttles in the constellation according to FIG. 1 in terms of the number of displacement spaces served and the shortness of the respective suction phases . (Assumption of the same speed and the same relative filling).

This has a favorable effect on the price and the manufacturing tolerances. In addition, a special contouring of the control cross section via the throttle opening path is rather possible, and the control principle can also be applied to extremely small pumps.

A second advantage is that due to the short suction phases and uniform phase shift of the displacement movement (= control by the rotary shaft with eccentrics), there is relatively little or no overlap of the suction phases. (There is no overlap of the suction phases if the height of the opening 35 is kept so small that the swept area angular range of the eccentric 11 or the rotary shaft 12 during the opening of the opening 35 by the piston 9 is a maximum of 360 ° / number of displacement elements is).

This is equivalent to connecting one and the same choke in succession to the different displacement elements. This means that the throttle cross-section is the same for each displacement element as an ideal prerequisite for equal filling or delivery of all displacement elements. A third advantage results if the described release angle is a certain amount smaller than the 360 ° number of displacement elements. This results in more or less short intermediate phases in which none of the displacement spaces sucks.

the filling of the channel pieces 36, 36a, 36b between the respective 2/2-way valve 21, 21a, 21b and the respective inlet cross-sections 35 (35a, 35b hidden, not visible in the drawing) can basically continue between the suction phases. This also helps in the channel pieces 36, 36a, 36b, i.e. H. to reach at least one cavity void up to the displacement space limit in the form of the inlet cross-section 35.

In the intermediate phases, the pressure p ₃ in the connecting channels can even rise to a maximum of p ₂ , since no fluid is removed from the channel pieces 36, 36a, 36b by any suction element. This leads to a temporary larger opening of the 2/2-way valves and to an acceleration of the filling of the duct sections.

FIG. 3 shows a particularly favorable embodiment of the control device of FIG. 1.

The possibility achieved due to the absence of cavities is shown here to arrange the adjusting device for the throttling 2/2-way valves 21 integrated here in the pump at a greater distance from these or the individual displacement spaces. This allows several or all control elements to be combined to form an actuator 60 in the form of a continuous directional valve with only one drive, which in turn then makes simple manual operation possible, for example. In the case of electrical pump adjustment, the need for only one converter for several or all displacement rooms is one great cost and space advantage.

Only the merging of the individual throttles belonging to the displacement elements into the continuous directional control valve 60 also permits optimal equal control of the individual throttles. As is known, the control openings of control slides and housings of such valves are usually manufactured in one setting, which means that these openings are positioned with little error and immovable relative to one another.

An important property of the invention is that the liquid volumes enclosed between the one adjusting device 27 and the individual throttling 2/2-way valves 21 in a channel are hardly elastic due to the lack of voids, so that hardly any additional liquid quantities enter or leave must flow out in order to achieve the respective steady state of a filling process or the time period between two filling processes. This means that the geometrical channel volumes may deviate greatly from one another, which is why the invention is suitable for all geometrical displacement arrangements (e.g. axial, radial, row for piston pumps). For all of these displacement arrangements, a location for the adjusting device 27 that is favorable in terms of installation space and appearance can be found.

In this example, the adjusting device 27 is even connected to the pump by hose lines 41, 41a, 41b, which allows the pump to be remotely controlled over a multiple length of the characteristic pump dimension (e.g. diameter in the case of a radial piston pump).

FIG. 3 also shows a further possible and advantageous variant of the invention in that an additional damper is the inherent one described above under FIG. 1 Damping added. The damper shown is only one example of possible designs. In this example, the respective throttling 2/2-way valves 21 which are actuated by pressure difference are connected to respective damping pistons 73 which can be moved back and forth in respective cylinders 70 in accordance with the movement of the slide of the 2/2-way valves 21. The effect of the damping is good and constant due to the lack of voids. Damping chambers 71 and 72 are formed in the respective cylinders 70 on opposite sides of the respective damping pistons 73. When the damping pistons 73 are shifted in accordance with the opening or closing of the respective slide of the associated 2/2-way valve 21, liquid flows past the piston from the chamber 71 into the chamber 72 or from the chamber 72 into the chamber 71 and through the guide gap Rod 74 and dampens the movement of the piston and therefore the corresponding slide of the 2/2-way valve 21. This helps to avoid an uncontrolled overshoot of the valve movement, as this would affect the flow core lines.

Fig. 3 also shows an inexpensive variant of the invention in that the pressure-differential actuated 2/2-way valves 21 are simultaneously designed as inlet valves, which saves effort.

4 and 5 show in cross section or in longitudinal section a particularly favorable embodiment of a pump with a control device according to the invention. The pump according to FIGS. 4 and 5 is equipped with four displacement spaces 129a-d, which are arranged in pairs above and below the drive shaft 110. The displacement space 129b cannot be seen in the drawing, since in FIG. 5 it lies behind the cutting plane (VV in FIG. 4) in the upper part of the drawing. A respective piston or displacer 117 is provided for each displacement space. The displacers 117 are held in contact by two springs 135 in contact with two drive rings 114 eccentrically mounted on the drive shaft 110. The drive rings 114 are rotatably supported by means of needle bearings 115 on eccentrics 113 which are connected to the drive shaft 110 in a rotationally fixed manner relative to one another.

The springs 135 for the respective displacement pistons 117 are supported on a plate-shaped abutment 116 at the end of each individual displacement piston and the drive ring 114 presses on the respective sides of the spring abutments 116 opposite the displacement piston 117. The rotation of the drive shaft 110 therefore causes them to rotate with it non-rotatably connected eccentric 113 and the rings 114 a reciprocating movement of the displacement piston 117, the lifting movement of the upper displacement piston 117 being offset by 180 ° to the lifting movement of the respectively opposite lower displacement piston 117. This means, for example, that the displacement space 129a has its smallest volume, while the displacement space 129b has its largest volume and vice versa. The two eccentrics 113 are connected to the rotary shaft 110 offset from one another by 90 °, so that the stroke phase difference of two displacement pistons 117 arranged next to one another, i.e. of the lower displacement piston 117 in FIG. 5 and the upper displacement piston is also 90 °. On the one hand, this contributes to a smooth running of the pump, and on the other hand it contributes to an even supply of liquid.

The rotary shaft 110 is rotatably supported in the main housing 138 of the pump via the ball bearing 136 and the roller bearing 137.

For each displacement space 129a-d (of which the displacement space 129c is not shown), a respective inlet valve 134 and a respective outlet valve 118 are provided. The respective inlet and outlet valve pairs 134, 118, which belong to the respective displacement spaces 129a-d, are accommodated in respective housing parts 133a-133d, in which the cylinders forming the displacement spaces 129a-d and serving to hold the displacement pistons 117 are also arranged are. These housing parts 133a-d each have a cylindrical extension which is arranged coaxially with the respective cylinder, ie with the respective displacement piston 117, and is inserted in a corresponding cylinder bore in the main housing part 138. There is a respective ring seal between the cylindrical extension of each housing part 133a-d and the housing 138, so that the main housing 138 is sealed against leakage. The cylindrical extension of each housing part 133a-d also has an annular shoulder on which the end of the respective spring 135 facing away from the plate-shaped abutment 116 is supported. That is, the ring shoulder forms a further abutment for the spring 135.

Each housing part 133a-133d is also provided with a respective valve cover 119a-d, the individual valve covers 119a-d having a respective cylindrical recess 121 which is arranged coaxially with the cylindrical extension of the respectively assigned housing part 133a-d and a stem part of the Intake valve 134 and the components cooperating therewith, which are shown in FIGS. 6A and 6B on an enlarged scale. The valve covers 119a-d and the housing parts 133a-d are screwed to the crankcase 138 by means of continuous screws, which are shown in FIG. 5.

On the left side of FIG. 4 one can see a hollow rotary slide valve 150 integrated into the construction, which is designed, for example, in accordance with DE-PS 37 14 691 can be. In this embodiment, the valve 150 is the adjustable element which is used to control the pressure-differential-operated, throttling 2/2-way valves, which in this embodiment are formed by the respective inlet valves 134 with the associated parts, as will be described in more detail later .

Starting from the rotary slide valve 150, four distributor bores or distribution paths 130a-d are provided (130c not shown), which lead to the respective inlet valves 134, each in a chamber 134a-d on the stem side of the valve, immediately adjacent to the respective one Ventil¬ seat, the chamber 134c is not shown. Starting from each distributor path 130a-d, there are respective oblique bores 127a-d in the respective cylinder heads 119a-d, which open into the cylindrical spaces 121, the oblique bores 127c and 127d not being shown.

On the input side, the hollow rotary slide valve 150, which in this example is designed as an easily replaceable cartridge, receives liquid in the direction of arrow E via a housing bore 132 from a reservoir 1 with the pressure p2, as shown for example in FIG. 3. The fluid reaches the interior of the hollow rotary valve without a significant pressure loss via a constantly open, sufficiently large inlet cross-section 156. By rotating the hollow rotary valve, this is done by means of an electric drive 158 (FIG. 5) or a gas linkage, which is not in itself is shown, but engages on part 159, can be achieved by the interaction of elongated linear control slots 155a-155d in the hollow rotary valve 150 with the mouth edges of the distributor bores 130a-d (130c not shown), so that the pressures p3 prevailing in the distribution lines 130a-d by means of the adjusting element 159 exactly and can be set quickly.

In particular, the valve cartridge on the rear side, not shown, in each chamber can have symmetrically opposite identical openings 155a-155d and 156 and the movable slide can be made very thin-walled, so that the valve has the advantages of a valve according to DE-PS 37 14 691.

As can be read in DE-PS 37 14 691, rotary valves or axial slide valves of this type have the advantage that they can be actuated very quickly with low actuating forces due to low friction, low inertia and low flow forces, so that the electric actuator (actuator motor ) 158 can be small and inexpensive. On the output side, as provided in the previous embodiments, drain holes 112a-d depart from the respective outlet valves 118, of which the flow holes 112c and d are not shown, which merge into a common drain line 111 which, for example, leads to the "common rail" of one Common rail diesel injection system leads.

The pressure p3 in the distributor lines 130a to 13Od is communicated via the oblique bores 127a-d in the respective cylinder spaces 121 and here acts on the valve 134 via the cross-sectional area of the stem of the valve 134 in the opening direction. In the closed state of the valve 134, the same pressure p3 also acts on the side of the valve head facing the chamber 134 in the opening direction of the valve. The two springs 125 and 126 exert a closing force on the valve 124 at this stage. The relatively strong spring 125, which acts on the abutment 124 at the end of the valve stem, permanently exerts a closing force on the valve 124, while the relatively weak spring 126 is supported on a spring plate 126T which is arranged displaceably with respect to the valve 124 in the chamber 121. In the closed state of the valve and when the spring plate 126T is in contact with the abutment 124, the spring 126 also exerts a closing force on the valve 134. The spring plate 126T with spring 126 is primarily used for damping purposes. When the respective displacement chamber 129a-d is enlarged by moving the respective displacement piston away from the top dead center (OTP), there is a lower pressure on the displacement chamber side of the valve than in the cylindrical chamber 121, so that a total force acts on the valve member 134a that leads to an opening of this link. Both the strong spring 125 and the weak spring 126 are compressed. The liquid located below the spring plate 126T escapes through the damping openings in the spring plate 126T and therefore slows down the opening of the valve member 134.

The height of the opening stroke of the valve member 134 and the amount of liquid that flows past the head of the valve member 134 into the displacement chamber 129 depends on the pressure p3 in the distributor line 130.

During the displacement movement of the displacer piston 117, the volume of the displacer chamber 129 decreases and the pressure in this chamber increases, albeit initially only slightly, owing to the small amount of gas or liquid molecules escaping. On the one hand, this leads to a closing force being exerted on the valve member 134 which is greater than the opening force, so that the valve 134 closes. At this stage, the damping openings in spring plate 126T operate to dampen the closing movement of the spring plate, so that valve 134 closes the valve seat relatively gently and spring plate 126T also gently engages abutment 124 at a later time. This means that the damper is so is that it is only effective during the opening stroke of the throttling valve, that is, in the phase in which vibrations are most likely to be initiated and would be effective for the longest. 6, the damping piston can remain behind the valve movement in the closing phase. Through the opening that is released, fluid flows into the damper chamber under the damping piston and prevents the creation of negative pressure and cavities. The increasing pressure in the displacement spaces 129a-d also causes the respective exhaust valves 118 to lift off, so that diesel enters the lines 112a-d and 111 with the desired outlet pressure.

This arrangement has several advantages. The valve 150 can be integrated into the pump construction in a space-saving manner, since it does not matter on differently long distribution paths 130a-d. The design of the valve 150 with elongated linear slots 155a-d permits particularly good controllability of the pump down to the smallest delivery rates.

The use of poppet valves 134a-d as inlet valves, which also serve here as the pressure difference-actuated, throttling 2 / w-way valves, is generally the more economical variant than the use of slide valves. Above all, the displacement space has less leakage path, which is particularly important for pumps for the highest pressures, low speeds and lowest viscosities (as they occur in connection with common rail diesel injection) if the highest levels of efficiency are to be achieved. The tightness of the inlet seat valves 134 also has a positive effect on the equal conveyance from displacement space 129a-d to displacement space 129a-d, since leakage is generally subject to component tolerance. The general delivery characteristics of the pump can also be better adhered to in series production in designs with a seat valve. Vibrations of the throttling valves can - like general vibrations - lead to spring breakage or, in the case of seat valves, increased wear or shaft breakage, here these vibrations damage above all with regard to the conveying characteristic, which is thereby changed. Vibrations often occur randomly as a result of stochastically fluctuating damping effects or excitations. In such a case, stochastic flow fluctuations or hysteresis effects would occur on the pump, both of which would make the use of the pumps for control purposes more difficult. The use of a damper on the throttling valve is therefore proposed for the purpose of defined valve damping. With simple piston dampers of known design, the damping forces also generate negative pressures, which in turn can create cavities which are harmful to the damping function. This can be remedied by using a larger valve preload when using such a damper. If the damping piston diameter is kept large, roughly the size of a valve diameter, the negative pressures and the required additional valve preloads are reduced. This is desirable because the form of pumps should be kept as low as possible.

The possibility of arranging the adjusting elements at a greater distance from the throttling valves or individual displacement spaces allows several or all adjusting elements to be combined to form an actuator with only one drive, which in turn then enables, for example, simple manual actuation. In the case of electrical pump adjustment, the need for only one converter for several or all displacement spaces is a great advantage in terms of cost and installation space. In addition, the liquid volumes enclosed between an adjusting element and a throttling valve in a channel are hardly elastic due to the lack of cavities, so that hardly any additional liquid quantity has to flow in or out in order to determine the respective steady state of a filling process or between two Filling processes are to reach the time period. Thus, the geometrical channel volumes may deviate greatly from one another, which is why the invention is suitable for all geometrical displacement arrangements (for example axial, radial, row in the case of piston pumps) and for all of these a location for the adjusting device 27 which is favorable in terms of installation space and appearance can be found.

7 now shows, for throttle actuating elements, some special features of the design of the throttling valves, for example the valves 30 in FIG. 1 or 150 in the embodiment according to FIGS. 4 to 6.

1, 3, 4, 5 each with a throttle point. per displacement space, however, this is finite.

With the arrangement according to the invention, the pressure difference at the adjusting element enters the metered amount of liquid with the root of the pressure difference. With a fixed feed pressure, however, this pressure difference decreases with increasing throttle valve opening. The use of a differential pressure valve 40 in FIG. 2 shows how this pressure difference can in principle be kept constant by using the differential pressure valve to change the admission pressure parallel to the pressure upstream of the throttling valve.

However, the same goal can at least be largely achieved in that the spring-loaded throttling 2/2 way valves have a steep opening characteristic, which is achieved by a soft spring or a large pressurized valve surface or a combination of both, and in that the feed pressure p2 is sufficiently high so that even for maximum pump volume flow, ie large valve Opening, the pressure difference across the adjustment device is not significantly reduced. These measures basically ensure that the flows at the throttle elements are only slightly influenced by variations in the spring stiffness or spring preload of the inlet valve springs or by differences in the effective valve area. This design also eliminates the need for precise spring sorting or the setting of the spring preload on each individual inlet valve.

8, 9 and 10 show schematically for the designs according to FIGS. 3 and 4, 5 different displacement space fillings at the same speed and as with the spring design, as explained above, the dynamic process of a working cycle takes place. FIG. 8 shows the state for the full filling or conveying of the displacer spaces 15, FIG. 9 the state for half the filling or conveying of the displacer spaces 15 and FIG. 10 the state of not exactly zero conveying of the displacer spaces 15 and as a function of the angle of rotation of the drive shaft, based on the top dead center OTP and the bottom dead center UTP of the respective displacement _pistons 9. The valve cross-sectional _profiles A _valve and the pressure p5 in the cylinder during the suction stroke are V = V _maχ and for in FIGS. 8 and 9 for full displacement _{space filling} Half the displacement _{space filling} V = 0.5 V _is almost rectangular and the pressure di erences P _Speιse - P ^ _a ^ _g = p2 - p3 during filling are stable despite the dynamics and almost the same for all flow rates.

The reason that the stroke of the throttling 2/2 way valve 21 and the pressure p _{4 can be set} almost immediately and stably is due to the fact that according to the invention in channels 31a, b, c or 31 a, b, c or 130 a, b, c, d or adjusting element 27 respectively. 150 avoided cavitation. For the pressure p _s in the displacement chamber 15, as already explained above, a constant low value, in the example close to zero, is established beyond UTP until the valve closes.

In this way, quasi-constant boundary conditions prevail in the form of quasi-constant values of the feed pressure p ₂ and the displacement space pressure p ₅ during the entire suction process (valve opening to closing).

Because of the freedom or lack of voids achieved in connections 41a, b, c and 130a, b, c, d according to the invention, incompressibility may be assumed between inflow with pressure p ₂ and displacement space p ₅ . Thus, the pressure p ₃ upstream of the throttling valve 21 occurs without significant delay due to the continuity condition that the flow at the individual throttle cross-section 30, V _{30 must be} equal to r the flow at the throttling valve 21, V ₂₁ :

c _λ , c ₂ constants that define A ₂₁ (p ₃ ) and the liquid density.

A specific A ^ is thus assigned a specific A ^ (p ₃ ) in the suction phase and also a specific p ₃ via the constants c ₁ and c ₂ .

Compliance with the fixed assignment, e.g. B. as security against overshooting the valve 21 when opening, the inherent damping of the arrangement according to the invention already described above and the additional damping 70 or the damping of the valve described in FIG. 6A.

The pressures p ₃ for the different cases of filling are close to each other due to the special valve design, see FIG. 7, in comparison to p ₂ and p ₅ . This simplifies the above equation

When the displacement space 15 is completely filled according to FIG. 8, the free flow cross section A _valve through the inlet valves 28 assumes the maximum value. When the displacement space 15 is half filled according to FIG. 9, the valves 28 are only partially open. The delivered volumes V correspond to the area under the volume flow functions. 10 shows the case in which the delivery is currently 0 and the filling therefore also tends towards O. In Fig. 10 as a borderline case, the liquid is still compressed and decompressed to p ₅ = P ₆ at top dead center, that is to say the high pressure of the system, but nothing is pushed out in between. Despite delivery O, the displacers can be filled slightly to cover any piston leakage as a result of the compression / decompression. A minimal opening A _s (V → O) is therefore shown in FIG. 10. The duration of this opening extends approximately over the entire revolution, interrupted only by the relatively short compression / decompression phase. Similar courses are created for the further embodiments according to FIGS. 2, 3 and 4 to 6 and 13, 14, 15, 16 and 17.

Fig. 11 shows the flow rate characteristic volume flow V = dV / dt as a function of to

D _{rossel «} ^ ^er adjusting _throttle 30 of this control device. The characteristic curves increase asymptotically from the respective _{limit speed} C _ra _limit i ^{to C} _{m limit} against twice the volume flow of the limit speed, since the suction time is also influential in addition to the suction cross-section and this varies with the delivery rate per stroke towards zero from originally half a turn to almost to a full revolution, ie increased twice, as can also be seen from FIGS. 8, 9, 10.

FIG. 12 shows the corresponding flow characteristics for slot-controlled pumps, as in the embodiment according to FIG. 2.

FIG. 13 shows an embodiment similar to FIG. 3, but with a different design of the throttling 2/2-way valves actuated by pressure difference and with a different type of actuating throttle actuation. The 2/2-way valves of the embodiment according to FIG. 13 each consist of a ball 54, which is pressed against a valve seat by means of a spring 53. The movement of the ball 54 with respect to the valve seat when the valve is open depends on the pressure p3 prevailing in the respective line 31, 31a and 31b, as a result of which the filling of the displacement spaces is controlled as a function of p3. 1 is particularly suitable for integration in analog control loops, a switching valve 50 as a converter according to FIG. 13 has advantages in connection with digital electronics.

The illustration in FIG. 13 shows such an arrangement, wherein, as in FIG. 2, with slot-controlled pumps and with an opening angle adapted to the number of cylinders, one switching valve 50 is sufficient for several displacement elements 9. 14 shows an embodiment in which only one 2/2-way valve 81 is used for three displacement spaces in this example, the 2/2-way valve 81 being arranged outside the pump and the individual displacement spaces 15 feeds via lines 36, 36a and 36b.

In this example, the adjusting device consists of an adjustable displacement machine 84 which has a flow-limiting function. However, the displacement machine 84 is preferably driven by an electric machine which is variable in speed. The displacement machine is designed as a constant displacement machine and draws the liquid to be conveyed from the line 33 directly or indirectly via the system 7 from the liquid reservoir. The pressure relief valve 80 in this case functions as a safety valve or relief valve. This prevents an inadmissible increase in the pressure difference at the pre-feed pump if it is adjusted to a position in which it delivers more than the maximum amount of swallowing by the main pump.

If the throttling 2/2-way valve in this example was not shown as a spring-loaded check valve 81, but rather similar to valve 21 in FIG. 2, ie as a slide valve without deliberate reaction of the pressure p ₄ on the valve opening, then a certain valve opening would still exist if there are interruptions between the suction phases of the individual displacement elements. In such interruption phases, the pressure p ₄ quickly increases to the pressure p ₃ , as a result of which the void fraction in the channels 7, 7a, 7b is reduced in each case. Such interruption phases are achieved by selecting the height of the opening 35 so that it is opened by the piston 9 for in each case only less than 360 ° / number of displacement spaces. FIG. 15 is very similar to the embodiment of FIG. 14 and also uses the presence of an adjustable pre-feed pump, here in the form of the adjustable displacement machine 86, which is driven at the pump speed or a speed proportional to this. In principle, for example, the drive of the displacement machine 84 can be accomplished via the drive shaft 12 of the pump.

FIG. 16 shows an embodiment similar to the embodiment of FIG. 3, in which the pre-feed pump 34 runs at constant speed, but in which the input pressure is controlled by controlling the spring preload of the pressure limiting valve 90, i.e. the variable pressure relief valve represents the adjustment device.

17 shows a solution which allows the liquid to be separated from the reservoir and the actuating medium (but the same fluid is also conceivable).

This constellation has advantages if the fluid to be pumped is very viscous or contains contaminants that could impair its function (example: common rail injection system for heavy oil engines), or if the variable pump is to be self-priming or only a very low admission pressure is to be used . All that is then required is a considerably less powerful pressure source 100 for the actuating fluid, such a pressure source often already being available (e.g. compressed air network).

The actuating fluid is conducted via lines 101 with the controllable pressure p10 to the individual 2/2-way valves 103. In this case, the pressure plO acts on the active surface 102 on one side of the slide of the 2/2-way valve 103, while a spring 104 and the output pressure of the 2/2-way valve via line 106 acts on the active surface 105 on the other side of the slide 102.

18 shows a schematic illustration of a pump in a radial design with three displacement pistons 9, only the central part of the pump housing around the drive shaft 12 being shown and only the upper displacement piston 9 being shown completely.

As can be seen, all three displacement pistons 9 work with a common eccentric cam 11, which rotates with the shaft 12.

As can be seen from the illustration of the upper displacement piston, the latter, like the two further displacement pistons, is always kept in contact with the eccentric cam 11 by means of a spring 200. Although in the present drawing all three displacement pistons are driven by the common eccentric cam 11, it would also be conceivable to move the displacement pistons in the axial direction of the drive shaft and to drive them via separate eccentric cams. Any other number of displacement pistons can also be selected.

It is essential in the filling degree variable displacement pump of FIG. 18 that the liquid to be displaced reaches the individual displacement pistons 9 via the interior and 202 of the pump housing.

As before, the connecting line to the liquid reservoir is provided with the reference number 33. The reference numeral 30 indicates an adjustable throttle element which leads into the interior 202 via the line 31. The pump in FIG. 18 is slot-controlled and has inlet slots 35 for this purpose (only shown for the upper displacement piston). wherein the inlet slots 35 communicate via a pressure-differential operated and throttling 2/2-way valve 51 (as shown in FIG. 13) and corresponding line sections 204 and 206 in the pump housing with the interior 202. As before, the reference numeral 17 indicates the outlet valve, which is connected via a line 18 to corresponding lines of the further displacement pistons 9 (not shown) and finally leads to the "common rail" of the internal combustion engine connected to it.

In order to secure the slot control over a corresponding angle of rotation range of the eccentric cam 11, an opening 208 is provided in the displacement piston 9, which communicates with the inlet slot 35 and cooperates with it in the desired angle range.

During operation of the pump, the individual displacement pistons 9 are moved back and forth by the eccentric cam 11 with the cooperation of the corresponding spring 200 in the respective cylinders 210. As a result, the fuel through the line 33, the throttle 30, the line 31, the interior 202, the line 206, the 2/2-way valve 51, the line 204, the inlet slot 35 and the opening 208 of the displacer 9 in the displacement space sucked and then flows out through the outlet valve 17 under the action of the displacer 9.

Due to the relatively large volume of the inner space 202, the invention succeeds in limiting the escape of gases in this inner space in such a way that the pump functions properly. It should also be pointed out that in this embodiment it is also conceivable to use the 2/2 way valves in the respective displacement piston 9 (not shown). literature

(1) Welschof, B.

Analytical studies on the possible use of a suction-throttled hydraulic pump for power control using the example of a hydrostatic auxiliary drive in a motor vehicle Dissertation RWTH Aachen, 1992

(2) Schneider, W.

Pumps for future diesel injection systems 0 + P (oil hydraulics and pneumatics) 36 (1992) 5, pp. 304-310

(3) Cooper Bessemer Fuel Injection Manual

(4) Schneider, W., Stöckli, M., Lutz T., Eberle, M. High pressure injection and exhaust gas recirculation MTZ 54 (1993) 11

(5) Schweitzer, P.H., Szebehely, V.G.

Gas Evolution in Liquids and Cavitation Journal of Applied Physics 21 (1950) Dec, S.1218-1224

(6) Fassbender, A.

Suction throttling - the influence of pressure medium and

temperature

0 + P (oil hydraulics and pneumatics) 37 (1993) 9

(7) Schmitt, Th.

Investigations of the steady and unsteady flow through throttle cross sections in fuel injection systems of diesel engines Diss. TU Munich, 1966

(8) Controls and regulations in heating and ventilation technology,

Impulse program for home automation 1986 Published by: Federal Office for Economic Affairs, 3003 Bern, Switzerland

Claims

Control device for a filling level variable pump with at least one displacement chamber, which works according to the suction throttle principle with an inevitable change in volume of the displacement chamber or the displacement chambers (15; 129a-d), and which the liquid (2) to be delivered from a liquid reservoir (1) with a free surface which is pressurized by a gas pressure (pl), usually atmospheric pressure, through a line (33), optionally via a hydraulic system (7), characterized in that on the low pressure side in front of the pump or in front of the displacement space or at least one pressure difference-operated throttling 2/2-way valve (21, 21a, 21b; 134a-d; 51-54; 81; 103) is arranged in the displacement spaces of the pump, or in front of one or more group (s) of displacement spaces of the pump, whose opening cross-section can be controlled via a pressure (p3, p3a, p3b; plO) applied to an active surface to adjust the degree of filling.

Control device according to claim 1, characterized in that, in the case of a liquid reservoir with a pressure (pl) of one bar absolute, the pressure in front of the throttling 2/2 directional control valve actuated by the pressure difference, at which the latter opens, ie the pressure (p3, p3a , p3b), is at least about 0.9 bar absolute and is preferably in the range from 1.0 to 1.5 bar absolute, a higher pressure also being permissible or the pressure mentioned being chosen to be correspondingly higher at a higher reservoir pressure (pl), ie is at least 90% and preferably 100 to 150% of the reservoir pressure (pl), a minimum pressure difference of over 150% of pl also being permissible. Control device for a filling level variable pump according to claim 1, characterized in that the or each pressure difference-operated throttling 2/2-way valve (21, 21a, 21b; 134; 51, 52, 53, 54; 81) with increasing pressure (p3) on the inflow side, which bears against a first active surface (24) of the 2/2-way valve, opens and in the opposite direction on a second active surface (23) with the pressure (p4) or (p5) or with the pressure on the displacement chamber side the pressure (pl2) which is set in the return line (6) is acted upon near the tank pressure, and which continues to experience the force of a spring (22; 125; 53; 82) against the opening direction, the opening characteristic of the valve being shown in FIG is known to be designed via the spring preload and the pressure active surfaces in such a way that the valve only opens from or above a minimum _{opening pressure} difference _Δpmin between the active surface (24) and the outflow side, ie the pressure which is established (p4 or p5) et and that before each 2/2-way valve an adjusting device (27; 150) with a connection to the 2/2-way valve and is designed either as a throttling valve (30) or as a flow-regulating valve.

Control device for a filling degree variable pump according to claim 1, 2 or 3, characterized in that the adjusting device is preceded by a pressure source (8; 34) of known type for supplying with a sufficiently high pressure (p2), which supplies the liquid from the line (33) directly or indirectly via a hydraulic system from the liquid reservoir.

Control device according to one of the preceding claims, characterized in that the minimum opening pressure difference Δ _D .. to the active surfaces ((23 and 24) = p3-p4)), from which the 2/2-way valve or the 2/2-way valves open or open, the value

^Δ Pδmin = Pl ^_ πt ^ax [P _gas off ^{; P} steam from ^

oriented, whereby

^p c _as ^ ^he is _{out of} the smallest gas outlet pressure specific to the ness to be conveyed liquid-and standing in contact with it gaseous medium in the closed gehalte¬ NEN displacement chamber in the immediate pressure reduction generated by the Expan¬ sion movement of the displacer at maxima¬ ler mean displacement speed of the important operating range is measured.

P _{D f from} the lowest value of the vapor pressure

Liquid is in operation for the expected liquid temperature range, the pressure at which 10% of the mass can evaporate being selected in the case of a liquid mixture.

Control device according to claim 5, characterized gekennzeich¬ net that the minimum opening pressure difference Δ _pömιn to the effective areas ((23 and 24) = (p3-p4)) for pumps with the 2/2-way valves additionally arranged downstream spring-loaded check valve ( 28) the minimum opening _{pressure difference Δpömin is reduced} by the value of the opening pressure difference of the non-return _{inlet valves ΔPÖEV} , ie

^Δ Pόmin = P ^{1 "aX} t ^P Ga _S from '^{" P} Da _m pf au ^"Δ P6EV 7. Control device according to one of claims 2, 5 and 6, characterized in that the minimum opening pressure difference Δ _pδmin , from which the valve opens, zen both for pumps with inlet valves (35) and for pumps with inlet valves by a value Δ _{pötemp is} increased, which represents an additional pressure difference for safety in order to record the influence of the temperature dependence of the solubility coefficient, which has to be selected specifically for the liquid and depending on the operating conditions.

8. Control device according to one of the preceding claims, characterized in that at least one pressure-operated, throttling 2/2-way valve (21, 21a, 21b; 134; 51, 52, 55, 59; 81; 103) in front of groups of displacement spaces (15; 129a-d) or displacement space elements (16, 16a, 16b) or in front of individual displacement spaces (15; 129a-d).

9. Control device according to claim 8, characterized gekennzeich¬ net that the or each 2/2-way valve (21, 21a, 21b; 51, 52, 53, 54; 81; 103) very close to the inlet slots (35th ) or inlet valves (28) is arranged.

10. Control device according to one of the preceding claims, characterized in that in front of each of the 2/2-way valves (21, 21a, 21b; 134; 81) or a group of the 2/2-way valves (51, 52 , 53, 54; 103) or an adjustable element (27), each with a connection (31, 31a, 31b; 32; 41, 41a, 41b; 101) to the 2/2-way valves, is arranged in front of all these valves, wherein the adjustable elements form the adjusting device.

11. Control device according to one of the preceding Claims, characterized in that the adjusting device (27) designed as a throttling valve represents an electrically, mechanically, hydraulically or pneumatically adjustable throttling valve (30; 150).

12. Control device according to one of the preceding claims, characterized in that the adjusting device designed as a current regulating valve (27) is realized by the combination of a throttle valve (30) with a pressure differential valve (40).

13. Control device according to one of the preceding claims, characterized in that the adjusting device (27) is an electrically actuable, 2/2-way switching valve (60) which can be pulse-width modulated.

14. Control device according to claim 1, characterized gekennzeich¬ net that in front of each of the 2/2-way valves (81), or each a group of these valves or in front of all these valves an adjusting device (27) each with a connection (41) to the valves, the adjusting device (27) having a flow-restricting function, eg adjustable pump (86) driven with pump speed or a proportional speed, or a constant displacement machine (84) operated with a variable-speed electric machine (85), which directly or via the system (7) from the liquid reservoir.

15. Control device according to claim 1, characterized gekennzeich¬ net that the pressure in front of each of the 2/2-way valves (21, 21a, 21b) or each group of these valves or in front of all these valves by an electrically, mechanically or pneumatically adjustable, variable form Source, for example a constant pump (34) in combination with an adjustable pressure relief valve (90) can be controlled as an adjustable element, which draws the liquid from the line (33) directly or indirectly via the system (7) from the liquid reservoir (1) .

16. Control device for a filling level variable pump according to claim 1, characterized in that in front of the displacers (9) or in front of individual displacers (9), preferably very close to it, the opening cross section ADr of the or each spring-loaded, throttling 2nd / 2-way valve

(103) via a pressure (plO) applied to an active surface (102) of the 2/2-way valve from a pressure source

(100) a second fluid circuit, which may also use a different pressure medium, e.g. Air, contains, or is filled with the same liquid as in the liquid reservoir, is controllable.

17. Pump according to one of the preceding claims, characterized in that the throttling, spring-loaded 2/2-way valves are designed as inlet valves (124; 51, 52, 53, 54) of the pump, preferably as seat valves.

18. Pump according to one of the preceding claims, characterized in that the throttling, spring-loaded 2/2-way valves (21, 21a, 21b) act on damper (70).

19. Pump according to claim 17, characterized in that the damper (70) has a damping piston, and that the pressure difference Δ _{pδmin is} increased by half to the full amount of the maximum pressure difference occurring between the damper chambers (71) and (72) during operation . 20. Pump according to one of the preceding claims, characterized in that the adjustable, throttling valves (30, 30a, 30b; 60; 150) are each mechanically coupled to one another in groups or in their entirety or in an actuator with only one manual, mechanical , electric or pneumatic drive are summarized such that they are operated synchronously with each other.

21. Pump according to one of the preceding claims, characterized in that the spring-loaded, throttling 2/2-way valves (21, 21a, 21b; 124; 51, 52, 53, 54; 81; 103) have a steep opening characteristic, which by a soft spring (22; 52; 82; 104) or a large pressurized valve surface or a combination of both is achieved, and that the feed pressure (p2) is sufficiently high so that even for maximum pump volume flow, ie Large valve opening, the pressure difference over the adjusting device is not significantly reduced.

22. Pump according to claim 3, characterized in that the throttle cross-sectional areas of the adjustable throttles (30; 155a-d) represent an approximately linear function over the travel range, i.e. Represent slots.

23. Pump according to claim 20, characterized by the combination of the adjustable throttles to form an actuator in the form of a hollow slide valve, for example according to patent specification DE 37 14 691 C2 (150), in which a hollow, rotatable or axially displaceable slide body is arranged in a housing (138) provided with a plurality of chambers, wherein there are pairs and opposing openings (155a-d; 130a, 130b; 130c, 13Od) in the body (150) and housing (138), preferably such that they are opened or closed synchronously with one another.

24. Pump according to claim 23, characterized in that the opposite openings (155a-d; 130a, 130b; 130c, 13Od) are made in the body and / or housing (138) by means of wire erosion, wherein a threading hole is provided, which is in pump operation either not at all or only to provide a particularly large delivery rate, especially at very high speeds.

25. Control device according to claim 1, characterized gekenn¬ characterized in that the liquid supply to the filling degree variable pump via the, the displacement piston (9) actuating eccentric cam (11) containing the interior

(202) of the pump, which communicates with the displacement chambers via corresponding bores in the pump housing and inlet slots (35), the throttling 2/2 directional control valves (51) operated either in the pump housing in front of the inlet slots (35) or in the respective displacement pistons (9 ) are arranged.

26. Control device according to claim 25, characterized in that the adjusting device is an adjusting device common to the displacer, which is arranged in a line (33, 31) between the liquid reservoir and the interior in (202) of the pump housing.

27. Control device according to one of the preceding claims for a filling level variable pump with a plurality of cylinders and inlet slots (35), characterized gekenn¬ characterized in that the inlet slots (35) and the cooperating with them displacement piston (9) designed so are that the opening phase of the inlet slots is approximately 360 ° divided by the number of displacement pistons, preferably slightly less, whereby the control is accomplished with only one adjusting element.