KR100334520B1 - Control device for charge-regulating pump - Google Patents

Abstract

The control system of the charge-unregulated pump with at least one exclusion space is operated by the principle of suction-throttling by means of a positive change in the regenerative or static volumes, and in particular for the purpose of application to the common rail diesel engine injection system do. This allows accurate, precise, and highly dynamical control of the charge-unregulated pump at low cost without damaging the system by undesirable cavitation. At least one throttling 2/2-way valve (21, 21a, 21b; 134; 51, 52, 53, 54; 81; 103) is operated on the suction side of the pump by a pressure difference.

Such a 2/2-conversion valve may be used for some of the exclusion spaces or for the entire firm, or a corresponding valve of this type may be inserted in front of the respective exclusion spaces. The pressure differential control of one or each of the 2/2-way valves is performed by a regulator (27; 150) located on the inlet side of the 2/2-way valve and designed as a throttle valve or flow rate control valve.

Description

Control device for charge-regulating pump

The present invention relates to a control device for a charge-unregulated pump having at least one exhaust chamber which is actuated by a suction-throttling principle with a positive change in the volume of the exhaust chamber or of the exhaust chambers, , A fluid is delivered from a fluid tank having a free surface, which is subjected to atmospheric pressure, generally by gas pressure, by a hydraulic system without the supply of gas.

The charge-unregulated pump may be a lifting pump (e.g., a radial piston pump, an axial piston pump, a serial pump), or a rotating or pivoting piston pump (e.g., a vane-cell pump, Roller-cell pump). The present invention relates to such charge-unregulated pumps which serve as a principle of suction throttling using static exhaust motion. In these cases, partial filling of the exhaust chamber occurs as a result of controlled cavitation in the compressed fluid. Two pistons (vane-cell pumps, blocking-vane pumps, etc.) with reciprocating and rotary exhaust pistons can be regarded as exhaust pistons moving in the positive direction.

To increase energy efficiency in static pressure systems, there has long been a need for increased use of regulated pumps. However, the presently available design of such a regulating pump, which is usually manufactured by the principle of stroke adjustment, is either too expensive for many uses or too inefficient at partial delivery, i. E.

At the same time, the demand for electrical control of the direct cost savings of a regulated pump is increasing, because the combination of electronic and fluid technology tends to be continuous due to the gain / cost ratio of the non-decreasing electronics.

In order to be integrated into the tuning system (in the form of actuating member), future tuning systems must have special transfer-flow characteristics and have low hysteresis and sufficient speed (i.e., without having long idle times, for example) To reproduce them exactly. As is known, in order to operate a member in a control loop, this property is partly an essential and at least a significant advantage.

Furthermore, the high transfer uniformity of the individual exhaust pistons relative to each other is important due to noise generation on the one hand and consumption due to uniformity on the other, so any additional disturbances of different frequency differences that can stimulate the controller are also high pressure It is not accompanied by the system.

This type of static charge-unregulated pump can be used in a wide variety of applications such as vehicle, industrial, aviation and hydraulic machinery, and in particular common automotive hydraulic machinery and so-called common rail diesel injection systems. By using the phase-control principle in this charge-unregulated pump (see the list of references in the later section of this disclosure), very high efficiency can be achieved during partial transfer and especially for low viscosity media, very high pressure, Can be achieved. In particular, in a phase-controlled pump having reduced delivery per cycle, as compared to such a stroke-regulating pump, the duration of the pressure load of the exhaust piston body and associated work losses (such as piston- . This property of leak-freeness is particularly suitable for such pumps for common rail diesel injection technology, as well as for other reasons (cf. 2 and 4 of the reference list later in this disclosure).

Also, this is a low consumption of energy or power for adjustment. This is because it frequently occurs with the adjustment of the throttle at low pressure (US Pat. No. 4,907,949). This also enables manual adjustment, among other things.

In principle, the low power consumption also allows a very high adjustment dynamics, so that not only the necessary adjustments are computed quickly electronically, but also the adjustment can be performed with the use of high speed components for direct electric drive. Considering the small force, the size of the electric drive and the production cost are also lowered. In general, a small force makes it possible to adjust the hydraulic / mechanical system without substantially interacting between the adjustment parameter and the measurement signal.

An example of the high degree of coordination dynamics required is the common rail diesel injection system. The distribution line (common rail) and other volumes that carry the high pressure must be able to be pumped very quickly (by about 0.2 second when applied to the vehicle) at significantly higher pressures in response to signals from the engine electronics. For this purpose, the delivery of the pump should be able to be adjusted to a different range more quickly and is the minimum value at which a single pump operating cycle can be achieved. This can be read again in 4 of the list of references. In addition, when the pressure is even, this pump should be able to deliver different deliveries within the range of about two injections.

Other prior solutions, for example using separate control inlet valves, are particularly complex for pumps having a relatively large number of exhaust chambers. There is an important advantage in handling only one adjustment element for multiple cylinders.

An example of a typical control device for a charge-unregulated pump is known from PCT / EP 89/01057, which processes only one control device for a number of exhaust chambers, has a slot control device at the inlet side and is sufficient for many applications Do. The special flow guiding device in the eccentric housing results in a constant filling of all of the exhaust chambers and thus a constantly high delivery during some delivery which is low enough for many applications. However, the dynamics are not sufficient for various applications since all the cylinders are filled with the central eccentric housing and the latter must first be filled with the throttle element during the instant of transition to full delivery, And must be emptied before recovering from the transfer stream. However, the surface and gravity effects may then create a bubble accumulation at the sporadic site using a group, e. G., An escape that continues from the wall, which may still cause a variance and hysteresis effect of the statistical delivery during pump operation. For example, a state can be established, wherein some of the exhaust chambers accommodate more fluid and the other exhaust chamber forms a higher cavity during charging, and the transfer is thus not uniform.

The dynamics of the pump designed in a manner similar to that in the above-mentioned patent were measured by FaBbender (see reference 6 in the reference list). In this particular case, the transfer flow delays the movement of the motion member by an idle time of approximately 7 motion cycles. Therefore, the high-speed switching valve is insufficient. In the common rail diesel injection technology already adopted in the above example, the pressure of the distributor tube can be adjusted only hardly by the unacceptable rise due to its small volume in such timespan.

Another example of a common control device for pumps with inlet valves made of check valves is known from CH 674, 243 = EP-A-299,337. The disclosure of such techniques does not indicate any specific content in view of the pressure to be used. However, in this type of inspected pump, the pump is subject to cavitation during suction throttling, resulting in a considerable gas volume, which is very accurate and precisely and easily compromises simple control.

In consideration of cavitation during suction throttling using positive exhaust motions, the throttle-adjusting element is disposed near the exhaust chamber to achieve the high dynamics required. Thus, at least in a radial piston design, one adjustment device from each exhaust chamber or a complex mechanical linkage is also required. A single cylinder design is preferred and both cams or gears are proposed to achieve higher flow rates and higher pumping cycles. The limitations of the configuration and associated load-relationships are of particular concern, and this solution with n cams or n drive ratios for a pump with n cylinders results in a high cycle, high similarity of individual delivery trends, Also, as higher disturbances occur and the peak of the torque in the drive increases considerably (n times), the higher pressure rise in the cylinder results in a larger noise emission (n times), resulting in dangerous gas molecules from the cavity The re-entry process can no longer keep pace with the rate of pressure rise (see reference list 1) and cavitation losses can then occur under these conditions.

Cooperstown, Inc., of the United States "Oh Ha" Co., "Mount Vernon", manufactures a common type of twin cylinder piston pump for common rail sprayers, which has twin cylinders, The position of the regulating throttle element between the pair of cylinders is determined by the position of the exhaust piston arrangement (radial, axial flow, serial This pump is also provided with an inlet slot and the satisfactory tightness of the exhaust chamber is evidently achieved with a long stroke design (i.e. correspondingly large sealing length and smaller gap length) It requires an accurate crankshaft with cup tappets that absorb directional forces, Kinda visible.

In conclusion, the first disadvantage is that there must be a particularly favorable property when the pump is used in an adjustment system for optimum dynamics, accuracy of the transfer characteristic and lack of hysteresis, and each exhaust chamber must be equipped with its own operating element with a drive Or even that the operating element has to be connected to the central driving element with a complicated mechanism in addition to the problem corresponding to the equilibrium of the quantity. The compatibility of this goal between simplification (as little as possible or only one operating element) and high dynamics, accuracy of transfer characteristics and absence of hysteresis ensures that the individual exhaust chambers are far apart from one another, as in a radial or tandem arrangement When placed, or when there are multiple exhaust chambers. If the exhaust chamber is arranged in close proximity to one another as in an axial-piston pump, the centering of the adjusting element is basically possible, but the constituent space is too limited or installed in other parts.

These several reasons for limitation in the use of charge control by suction throttling with positive exhaust motion can be found in cavitation, which has hitherto been necessary to adjust the delivery and this is often desirable for the purpose of independence from viscosity, , It is generally already started in the throttling device and does not start only in the exhaust chamber.

Accordingly, it is an object of the present invention to provide a control device according to the preamble portion having general effectiveness for different pump types of exhaust type, which can be manufactured cost-effectively and at least significantly inhibits premature cavitation failure at low cost , Providing a much greater degree of freedom in practicing very interesting and aggressive transfer flow control. The provision of degrees of freedom is meaningful in terms of production cost, generally in terms of pump type, overall size and effective applicability generally described above for the design of the pump, and it is also possible to combine the operating elements and, for example, to operate them directly from the mechanical- It is possible to provide remote control possibilities by having the ability to place the adjustment element at any position on the pump without significant damage to the pump as well as having the ability to even position it slightly away from the pump.

Cavitation in fluids and associated cavitation damage in the case of fixed inflow volumes have been frequently studied in the past. However, the case of non-freezing and substantially non-freezing of cavitation in pump cylinders has been studied to a lesser extent so far. However, damage caused by cavity breakdown in materials common to pump construction is unexpected. One of several reasons is that the time is too short to effectively dissolve a substantial amount of gas or vapor. The Schweitzer in the bibliography (see 5) investigated the escape of dissolved gas from the fluid and found a diffusion time constant, which is the normal operating cycle period of the constant pressure pump. The pars-vendor (see bibliography 6) measured the gas-outlet pressure, which is actually very low for many suitable fluids.

The present invention is based on this physical phenomenon, and in the case of a gas in a gas atmosphere above a pressure p1, for example, a time-saturated fluid saturated in the atmosphere and a tank to be vented, Which is substantially better known to have a pronounced tendency when turbulence is additionally present during or around the < RTI ID = 0.0 > flow. ≪ / RTI > This is meaningless with respect to mass, but nevertheless it can fill a large part of the line or volume with respect to volume, requiring the filling or emptying operations described above until a new fixed state is established with respect to dynamics.

In order to achieve the above-mentioned object, the control device according to claim 1 or claim 15 is provided in the present invention. The main feature of the present invention lies in the fact that, due to the prior art connection of the pressurized manual throttle valve, the upstream of the individual exhaust chamber, the exhaust piston or the group of the whole pumps in accordance with the manner of the claims, the downstream pressure to the upstream point of the throttle valve _Ensures that it does not reach the pressure P1 of the fluid tank and preferably the amount DELTA _pTemp described after the addition of P1 and consequently limits the destructive cavitation to the exhaust chamber down to a relatively small volume of this valve . This procedure can be used to achieve a number of self-priming capabilities by not throttling all of the inlet valves, or by only slightly pressurizing them, in other respects, such as throttling with pressure loss in the pump, For example, kinks should be avoided as much as possible in order to reduce the risk of cavitation. For this purpose, the upstream pressure P3 of the throttle valve should be raised slightly in general by a pressure source of known type, for which a height difference between the fluid tank and the inlet valve is also made possible. For this reason, there is no obvious limit to the use or combination of the present invention as a result of these conditions, especially since most hydraulic systems, except for the hydraulic and fuel-supply systems of vehicles, operate in either case with pumps that occur at very low permissive pressures .

An important finding on which the present invention is based is that one liter of fuel or hydraulic fluid at atmospheric pressure can absorb almost 10% of the air volume in a dissolved state, which also enters the fuel tank of the vehicle. Therefore, at atmospheric pressure, a gas amount of approximately 100 cc is contained in 1 liter of fuel. When the pressure is reduced, this dissolved air is separated from the solution in gaseous form, and the amount of gas with respect to volume is expanded from 10 bar to 1000 cc, for example, from 0.1 bar according to the main negative pressure. This amount of gas very quickly fills the volume downstream of the valve to the exhaust chamber, greatly impairing the delivery and control of the fuel pump. The same object is also applied to other fluids. As much as 0.9 bar and preferably in the range between 1.0 and 1.5 , The delivery and control of the pump system is not affected by this since the formation of the gas volume is kept to a minimum or is completely suppressed.

The provisions of claims 5 and 6 take into account the particular properties of the fluid and the gas. The formula according to claim 5 and claim 6 is characterized in that for the pump with the inlet slot and for the pump with the automatic spring load inlet valve controlled by the exhaust piston movement both the typical or each 2 / The minimum open-pressure difference Can be determined. If the gas-outlet pressure (pgasout) and the vapor-outlet pressure (pvaporout) are unknown, then it is safe to set this pressure to 0 bar in the formula.

The so-called solubility parameter describes the dissolution reaction according to the Henry's equation, especially for fluids and gases:

cs = k * p

From here,

The saturated concentration of the dissolved gas or gas mixture in the cs fluid

p Pressure in the saturated equilibrium state (P1)

k = k (T) Solubility coefficient of a gas or gas mixture in a fluid.

In many systems, particularly for example in a vehicle, for example, from a low temperature tank to a high temperature engine, the delivered fluid may experience a sudden temperature change.

If the solubility coefficient decreases in the direction of the temperature change, a sudden supersaturation of the fluid occurs, which results in an already unobstructed gas release upstream of the throttle spring-load valve.

To ensure this, that is to maintain at least the saturation concentration, the maximum temperature-relationship reduction of the solubility coefficient k that occurs during operation can be prepared by increasing the minimum opening difference by the amount DELTA Ptemp.

If cs _x = cs ₁

At this time, using the Henry equation, the following equation is established.

k (Tx) px = k ( T 1) p 1, px / p 1 = k (T 1) / k (Tx)

Or _{△ ptemp = px - p 1 =} (px / p 1 - 1) p 1 = (k (T 1) / k (Tx) -1) p 1

When k (Tx) < k (T ₁ ), here

Tx and T ₁ is a fluid tank and the throttle spring-defines the maximum temperature difference of the fluid takes place in a time interval of several hours, during operation between the load valve.

The main advantage of the control device selected in accordance with the invention is that it can be regenerated at a low-hysteresis and low-idle time reaction as desired with the adjustment of the adjustment member. The precise calculation of the collecting member position and the pump flow rate is a control loop of the hydraulic system, especially in the case of a loop with strict conditions required for control dynamics, such as a common rail diesel system, Condition. In view of the theoretically infinitely fast set-up operation of the switching valve, the overall reaction associated with the delivery takes place already in the first consecutive complete inhalation action (which, in principle, never occurs quickly). Thus, in a hydraulic system in which the expected sudden change during consumption can be known, the delivery amount of the pump can also be changed at the same time.

With respect to the throttle valve near the exhaust chamber, a significant shortage of the cavity (and when the throttle valve is designed as an inlet valve with respect to the restriction of the exhaust chamber), the degree of freedom is obtained in most of the various pumps for the exhaust piston, Allows the use of various adjustment devices for the characteristics.

Particularly preferred embodiments of the control device according to the invention for charge-unregulated pumps can be obtained in other dependent claims.

Hereinafter, the present invention will be described in more detail by way of exemplary embodiments with reference to the drawings.

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

2 shows another variant of the control device according to the invention for a pump having an inlet slot controlled by an exhaust piston,

Figure 3 shows a special design of the control device according to the invention for the pump, the special spring characteristic and the regulating throttle to be connected to the inlet valve and the continuous direction valve designed with a damper,

4 shows a section of a finished pump with a control device according to the invention installed in the pump,

5 shows a schematic view of a pump having a control device of FIG. 4 with a partial longitudinal section taken along line V-V of FIG. 4,

Figures 6A and 6B show a diagram illustrating the mode of operation of the inlet valve of the pump for Figures 4 and 5, wherein Figure 6A shows the opening action and Figure 6B shows the closing action,

FIG. 7 shows a diagram for explaining the design characteristics of the throttle valve,

Figure 8 shows a graphical representation of the operating cycle of the pump according to Figure 3 for full transfer,

9 and 10 show the same representations for half transmission and " 0 " transmission, respectively,

11 shows the transfer-fluid characteristics of the pump according to FIG. 3,

12 shows a transmission-fluid characteristic similar to that of FIG. 11, but for a slot-controlled pump,

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

14 shows a design of a control device according to the invention for a charge-unregulated pump, wherein the adjustment device is formed by a variable displacement machine,

Figure 15 shows an embodiment similar to that of Figure 14,

Figure 16 shows another variant of the control device according to the invention for a charge-unregulated pump, in which the regulating pressure-relief valve acts as a regulating device,

Figure 17 shows a preferred variant of the control device according to the invention for a charge-non-regulating pump, in which the regulating device is operated with an auxiliary medium rather than a fluid to be pumped,

FIG. 18 shows a schematic view of another charge-unregulated pump according to the invention.

Figure 1 shows a first possible variant of a control device for a pump with an automatic action inlet valve.

The pump according to the schematic of Figure 1 has three individual exhaust pistons 9, only one of which is shown in the first figure. Three exhaust pistons are driven by the drive shaft 12 through respective eccentric cams 11 and each eccentric cam 11 is disposed in a lifting member 10 located at the lower end of the exhaust piston 9. [

In this case, the rotational motion A of the eccentric cam 11 causes a reciprocating motion B, and the exhaust piston 9 is moved to the two dead center positions C (lower dead center) and D (upper dead center) And acts as an exhaust piston reciprocating in the exhaust chamber between the exhaust chambers and causes periodic suction motions. As a result of the lifting member 10, the exhaust piston is not lifted apart from the eccentric cam 11 in any phase with respect to its movement (positive displacement). The inlet valve 28 and the outlet valve 17 are installed in a manner known per se for the respective exhaust chamber and the inlet valve 28 and the outlet valve 17 are connected to respective springs 28) in each case at the closed position by means of a valve. This means that the inlet valve 28 is designed as an inlet check valve. As a result of the movement of the exhaust piston 9, by the rotational movement of the eccentric cam 11, the inlet check valve is opened in a known manner by the generated pressure difference p4-p5, and a suction operation occurs. During the upward stroke of the exhaust piston 9, the amount of fluid collected so far is transferred from the exhaust chamber 15 through the outlet valve 17, that is, the latter is lifted from its seat against the action of the pressure spring, The fluid is passed through lines 18a and 18b along with a corresponding amount of fluid through line 18 to the common line 19 where pressure p6 is dominant and this is for example a common- common rail " injector of the " common rail " (distribution pipe).

As is common in this multi-piston arrangement, individual pistons or exhaust pistons 9 are used to achieve an equilibrium of the outlet pressure p6 in the cavity line and to ensure that the pump is operated with as little vibration as possible, . That is, if there are three exhaust pistons, as shown in the embodiment according to FIG. 1, the individual exhaust pistons perform lifting movements in each case with a phase shift of 120 DEG with respect to the adjacent exhaust pistons.

The amount of perforation through each exhaust piston is determined by the respective throttle-spring loaded 2/2-way valve 21 located upstream thereof and the regulating device 27 designed as the regulating throttle 30 in this example.

The adjusting device 27 is supplied to the common line 32 which supplies the diesel oil at the pressure p2, as in the case of the equally designed adjusting devices 27a and 27b. The diesel fuel 2 exits the fluid tank 1, which is in contact with the gas 3 at the pressure p1, i. E. Atmospheric air, at the interface 4. The fluid can be saturated with gas. The fluid first flows through the system 7, wherein no other gas flows into the fluid. Since the pressure is increased from pi to p2, the pressure increasing device, i. E. The pressure source 8 in this example, is incorporated into the system 7.

The diesel oil in the line 32 then flows through the three regulating throttles 30, 30a and 30b and the two / two-way throttle valves 21, 21a and 21b which are attached thereto and operated by pressure difference. By the continuity equation for the incompressible medium (which can only be assumed on the assumption that there is no cavity), the amount of perforation through each adjustment throttle and the 2/2-way throttle valve 21, 21a, and 21b attached thereto is the same Do. A pressure p3 at the working surface 24 of the 2/2-way throttle valve 21 and a pressure p12 at the other side of the 2/2-way throttle valve, such as a tank near the p1 of the working surface 23, And the force of the spring 22 due to the movement of the throttle are established. The regulating throttles 30, 30a, 30b can be adjusted individually to theoretically match each other.

Figure 1 shows a more important advantage of the present invention. The system has an intrinsic damping effect which is increased due to more abrupt throttling using the valve operating surfaces 24, 24a, 24b and adjustment throttle 30, 30a, 30b and which is important for maintenance and regeneration of the transfer characteristics 1 and 11). The operation surfaces 24, 24a, 24b and the distribution pipes 31, 31a, 31b are closed when there is only a slight overshooting of the 2/2-way throttle valve 21, 21a, damping function in that the increase in volume generated by multiplying the difference between the stroke of the sikindaneun significant amount △ p ₃ generates a lowering of the pressure p ₃ by the reaction with the overshooting, and this is due to lack of the cavity according to the invention .

The lower the throttle setting, the longer the time and the damping effect last longer until the medium can flow.

In this exemplary embodiment, the 2/2-way throttle valves 21, 21a and 21b operated on the pressure differential are connected to the return 6, respectively, on its working face 23, The pressure p12 acts on the operation surface 23.

The advantage of this arrangement is that the spring 22 can be chosen to be very weak depending on the size of the working surface 23 and the adjustment of the valve 21 against the opening pressure p3 acting on the other working surface 24 instead It is not much help for pressure than for. This is because there is already a substantial amount of required pressurization and further pressurization using the pressure p12 on the operating surface 23.

2 shows a control device similar to that of FIG. 1, the difference being that the pump has an inlet slot 35 and only one central regulating device 27 with an adjusting throttle 30 is installed. Pumps with inlet slots can generally be manufactured at a lower cost compared to having inlet valves, but this is used at somewhat higher pressure and lower viscosity pressure media.

In this exemplary embodiment, the low-cost objective is achieved with a central adjustment device 27 that basically allows simple manual or electrical adjustment. The individual regulating throttle 30 in the regulating device 27 can also be manufactured in a manner known per se and inexpensively. In the suction step, the pressure difference (p2 - p3) in the regulation throttling is kept substantially constant regardless of the amount of perfusion by the pressure difference valve connected in parallel, and as a result, the adjustment throttle 30 and the pressure difference valve 40 ) Brings about the effect of the flow control valve. A simple operation using the same adjustment throttle 30 for all the exhaust elements 16, 16a, 16b provides another advantage in this arrangement with inlet side slot control.

The first advantage is that, with respect to the characteristic rotational speed and the characteristics of the exhaust chamber, relative to the number of exhaust chambers and the respective short suction steps to be fed, the control section of the throttle 30 is arranged, for example, Which is substantially greater than in the individual throttle of FIG. (Assuming the same rotational speed and same relative charge).

This has an advantageous effect on the price and production tolerance drawing. Further, the special shape of the control section during the throttle opening movement becomes more easily possible, as is the case with the application of the control principle to extremely small pumps.

A second advantage is that the superimposition of the inhalation phase is relatively small or none at all due to the short inhalation phase of the exhaust motion and the constant phase shift (= exhaust control by the drive shaft with eccentric cam). The overlapping of the suction steps is such that the height of the inlet slot 35 remains very low so that the angle of the eccentric cam 11 or the drive shaft 12 during opening of the inlet slot 35 by the area of the cover, If the range is up to 360 ° / the number of exhaust elements, there will be no.

This is like locking the same throttle continuously on various exhaust elements. This means that the throttle cross section for each exhaust element is identical as an ideal precondition for the same charge or uniform delivery of all exhaust elements.

A third advantage is obtained when the described opening angle is somewhat less than 360 [deg.] / The number of exhaust elements. A somewhat short intermediate stage is obtained at which time none of the exhaust chambers is sucked.

36a, 36b between the respective 2/2-way throttle valve 21, 21a, 21b and each inlet slot 35 (obscured 35a, 35b can not be shown in the figure) The charging can basically continue between the suction steps. This also contributes to achieving at least a lack of cavity with respect to the exhaust chamber restriction in the channel portions 36, 36a, 36b, i.e. in the form of inlet slots 35. [

In the intermediate stage, the pressure p3 in the connection channel can rise to the maximum pressure p2 since no exhaust element can extract the fluid from the channel portions 36, 36a, 36b by the suction operation. This causes the 2/2-way throttle valve to open more instantaneously and accelerate the charging of the channel portion.

Figure 3 shows a particularly advantageous embodiment for the control device of Figure 1.

This shows the possibility of arranging the adjustment device for the 2/2-way throttle valve 21 which is obtained as a result of the absence of the cavity and which is incorporated in the pump farther away from this or the individual exhaust chamber. This makes it possible to couple a number or all of the operating elements to one pulse width modulating 2/2-way switching valve 60 in the form of a continuous directional valve with only one drive, At this time. In the case of an electric pump adjustment, there is a great advantage in terms of cost and structural space, since only one adjustment device is required for many or all of the exhaust chambers.

Simply coupling the individual throttle belonging to the exhaust element to the switching valve 60 also permits optimal equal control of the individual throttle. As is known, the control slides of such valves and the control orifices of the housing are generally made of a fastening device, which means the low defect fastening position of the orifice with respect to each other.

An important feature of the present invention is that the amount of fluid contained between one conditioning device 27 and the individual 2/2-way throttle valve 21 in the channel is virtually elastically deformed due to the absence of the cavity, Or very little to the inside or outside in order to achieve a particular fixed state of the time period located between the two charging operations. Thus, geometric channel volumes are allowed to vary considerably from one another, which is why the present invention is suitable for all geometric exhaust piston arrangements (for example, axial flow, radial and tandem in the case of a piston pump). The position of the adjustment device 27, which is advantageous in structural space and appearance, can be found in all such exhaust piston arrangements.

In this example, the regulating device 27 is connected to the pump by means of the distributor pipes 41, 41a, 41b so that the length of the pump over the length of the various characteristic pump dimensions (for example in the case of a radial piston pump) Allows possibility for remote control.

FIG. 3 also shows further possible and advantageous variants of the invention, as long as the additional damper supplies the inherent damping as described in FIG. The damper shown is only one example of a possible design. In this example, each 2/2-way throttle valve 21 operated by a pressure difference is connected to a respective damping piston 73, which, as the slide of the 2/2-way throttle valve 21 moves, (70). Since the cavity is small, the damping effect is good and constant. At the same time, damping chambers 71 and 72 are formed in the dampers 70 on the opposite sides of each damping piston 73. During the evacuation of the damping piston 73 in accordance with the opening or closing of each slide of the 2/2-way throttle valve 21, fluid flows from the damping chamber 71 past the piston to the damping chamber 72, Flows through the guide gap of the rod 74 from the piston 72 to the damping chamber 71 and weakens the movement of the piston and hence the movement of the corresponding slide of the 2/2-way throttle valve 21. This affects the transfer-fluid characteristics, thus preventing overshooting of uncontrolled valve movement.

3 also shows an advantageous variant of the invention in which the 2/2-way throttle valve 21 operated by pressure difference is designed as an inlet valve at the same time, resulting in cost savings.

4 and 5 show particularly advantageous designs of the pump with the control device according to the invention, in cross section and in longitudinal section, respectively. The pumps according to FIGS. 4 and 5 are provided with four exhaust chambers 129a-129d, which are arranged in pairs above and below the drive shaft 110. As shown in FIG. The exhaust chamber 129b is disposed behind the upper end face (V-V in FIG. 4) of the figure in FIG. 5 and therefore can not be shown in the drawing.

Each piston or exhaust piston 117 is installed in each exhaust chamber. The exhaust piston 117 is in contact with two drive rings 114 eccentrically mounted on the drive shaft 110 by respective springs 135. The drive ring 114 is rotatably mounted on the eccentric cam 113 by a needle bearing 115, which is fixedly connected to the drive shaft 110 in a manner that is offset relative to each other.

A spring 135 for each exhaust piston 117 is supported by a pedestal 116 such as a plate at the end of each respective exhaust piston and the drive ring 114 is connected to a spring pedestal 116 located on the opposite side of the exhaust piston ) On each side. The rotation of the drive shaft 110 causes the reciprocating motion of the exhaust piston 117 by the eccentric cam 113 and the drive ring 114 fixedly connected to the rotation of the drive shaft 110. The rotation of the drive shaft 110 causes the stroke of the upper exhaust piston 117 The movement occurs in such a way that it is offset 180 DEG apart with respect to the stroke movement of the respective opposite lower exhaust piston 117. This means, for example, that the exhaust chamber 129a has its minimum volume while the exhaust chamber 129b has its maximum volume. Since the two eccentric cams 113 are connected to the drive shaft 110 at intervals of 90 degrees with respect to each other, the two exhaust pistons 117 disposed adjacent to each other, that is, the lower exhaust pistons 117 and The stroke-phase difference of the top exhaust piston reaches 90 °. This on the one hand contributes to the static movement of the pump and on the other hand it contributes to the constant delivery of the fluid.

The drive shaft 110 is rotatably mounted in the main housing 138 of the pump by a ball bearing 136 and a roller bearing 137.

Each inlet valve 134 and each outlet valve 118 are installed in respective exhaust chambers 129a-129d (of which exhaust chamber 129c is not shown). Each pair of inlet valve 134 and outlet valve 118 belonging to each of the exhaust chambers 129a-129d are contained in respective housing portions 133a-133d, wherein the exhaust chambers 129a- And a cylinder for accommodating the exhaust piston 117 is also disposed. Each of these housing portions 133a-133d has a cylindrical extension coaxially disposed in each cylinder, i.e., each exhaust piston 117, which is inserted into a corresponding cylinder bore in the main housing portion 138. [ Each annular gasket is positioned between the cylindrical extensions of the respective housing portions 133a-133d and the housing 138 so that the main housing 138 is sealed against leakage. In addition, the cylindrical extension of each housing portion 133a-133d has an annular shoulder on which the end of each spring 135, which is directed to a pedestal 116 such as a plate, is supported. That is, the annular shoulder forms another pedestal for the spring 135.

Each of the housing portions 133a-133d is also provided in the respective valve covers 119a-119d, and each of the respective valve covers 119a-119d is coaxial with the cylindrical extension of the respective housing portion 133a-133d Which accommodates the protrusions of the inlet valve 134 and the components associated therewith and shown in enlarged dimensions in Figures 6A and 6B. The valve covers 119a-119d and the housing portions 133a-133d are threaded into the housing 138 by a continuous screw as shown in FIG.

On the left hand side of FIG. 4, a hollow rotary slide valve 150 is shown, which is integrated in the configuration and can be designed, for example, according to German Patent Specification 3,714,691. For this embodiment, the rotary slide valve 150 constitutes a regulating element, which serves to control the 2/2-way throttle valve operated by the pressure difference, which, in this embodiment, As well as inlet valve 134, respectively, with associated parts.

Starting from the rotary slide valve 150, four dispensing orifices or distributor tubes 130a-130d (130c not shown) are provided, which allow each inlet valve 134 to communicate with a respective valve Through the chamber 134a-134d on the projection side of the valve immediately adjacent to the seat, and the chamber 134c is not shown. Starting from each of the distribution pipes 130a-d, each of the valve covers 119a-119d is provided with respective inclined holes 127a-127d, which are open to the cylindrical space 121, 127d are shown.

On the inlet side, a hollow rotary slide valve 150 designed with a plug-in cartridge replaceable in a simple manner in this example is connected to the housing hole 132 (FIG. 3) from the tank 1 of pressure p2, for example, ) And receives the fluid in the direction of arrow E. The fluid is constantly open without significant pressure loss and passes through a sufficiently wide orifice 156 into the interior of the hollow rotating slide. The result of the rotation of the hollow rotary slide is generated by a gas link device which is electrically actuated drive 158 (FIG. 5) or not shown, but which engages operating element 159, 130d are provided with a combination of the linear control slots 155a-155d extending from the hollow rotary slide valve 150 having the mouse edges of the distribution lines 130a-130d (130c is not shown) p3 can be set accurately and quickly by operating element 159. [

In particular, the valve cartridge (not shown) on the backside can have the same orifices 115a-115d and 156 symmetrically facing each chamber, and the movable slide can be very thin in wall thickness, And has the advantage of a valve according to German Patent Specification No. 3,714,691.

The advantage of this type of rotating slide or shaft slide, as gathered from German Patent Specification No. 3,714,691, is that it can be operated very quickly and accurately by means of a small operating force as a result of low friction, low inertia and low oil- The operation drive section (operation drive motor) 158 can be manufactured in a compact and low-cost manner. Outlets 112a-112d extend from respective outlet valves 118, as outlined in the previous embodiment, and outflow holes 112c and 112d are connected to the common outlet line 111 ), Which leads to the "common rail" of a common rail diesel injection system, for example.

The pressure p3 in the distribution pipes 130a to 130d is transmitted through the inclined holes 127a-d of the respective cylindrical spaces 121 and passes through the end face of the projection of the inlet valve 134 Acting in the opening direction. In the closed state of the inlet valve 134, the same pressure p3 also acts in the valve opening direction on the side of the valve head toward the inlet valve 134. At this stage, the two springs 125 and 126 apply a sealing force to the pedestal 124. The relatively weak spring 126 permanently applies a hermetic force to the valve 124 while the relatively strong spring 125 engaging the pedestal 124 at the end of the valve projection permanently applies a sealing force to the valve 124, And is supported by a spring plate 126T which is arranged to be interchangeably disposed on one side. In the closed state of the valve, and when the spring plate 126T is supported on the pedestal 124, the spring 126 also exerts a sealing force on the inlet valve 134. [ However, the spring plate 126T with the spring 126 is primarily suitable for damping purposes. Since the low pressure is more prevalent on the exhaust chamber side of the valve than in the cylindrical space 121 when each of the exhaust chambers 129a to 129d is expanded as a result of movement of the respective exhaust pistons away from the upper dead center TDC, The force acts on the valve member 134a resulting in the opening of such a member. Simultaneously, both the strong spring 125 and the weak spring 126 are squeezed. The fluid located under the spring plate 126T exits through the damping orifice in the spring plate 126T and slows the opening of the inlet valve 134. [

The amount of open stalk of the inlet valve 134 and the amount of fluid flowing into the exhaust chamber 129 through the head of the inlet valve 134 depends on the pressure p3 of the distribution pipe 130. [

During the exhaust stroke of the exhaust piston 117, the volume of the exhaust chamber 129 is reduced and the pressure in such a space is increased, even though it is initially only slightly due to a small amount of gas or fluid molecules exiting. This result, on the other hand, is that the inlet valve 134 is sealed because a sealing force greater than the opening force is applied to the inlet valve 134. [ By this step, the damping orifice in the spring plate 126T acts to weaken the sealing movement of the spring plate, so that the inlet valve 134 closes relatively slowly with respect to the valve seat, and the spring plate 126T, And is gradually supported with respect to the base plate 124. This means that the damper is designed to be effective only during the open stroke of the throttle valve, that is, at the stage where the vibration is most easily generated and lasted the longest. According to FIG. 6, in the sealing step, the damping piston can be deployed later than the valve movement. The fluid flows through an orifice which is exposed to the damping space below the damping piston and which prevents negative pressure and cavities from being formed. In addition, the elevated pressure in the exhaust chambers 129a-129d causes each outlet valve 118 to also be raised, so that diesel oil passes through the line 112a-112d or 111 at the desired initial pressure.

This arrangement has various advantages. The valve 150 can be integrated into the pump configuration in a space-saving manner since it is not critical that it has different lengths of distribution tubes 130a-130d. The design of the valve 150 with the elongated linear slots 155a-155d allows for particularly good control of the pump at the bottom with very little transmission.

The use of the seat valve 134a-134d as an inlet valve here acting as a 2/2-way throttle valve operated in accordance with the present invention is generally a more cost effective variant than the use of a slide valve, The exhaust chamber has one less leakage path, which is particularly important for pumps of extreme high pressure, low rotational speed and low viscosity (as occurs in connection with common rail diesel injection) to achieve very high efficiencies. In addition, the tightening of the inlet valve 134 severely affects the same delivery from the exhaust chambers 129a-129d to the exhaust chambers 129a-d since the leaks are generally closely related to the component tolerances. The general transmission characteristics of the pump can be maintained more effectively in mass production in configurations with seat valves as well. Vibration of throttle valves can cause spring cracks like ordinary vibrations, increased wear or protruding cracks in the case of seat valves, where such vibrations are also harmful to transmission characteristics that vary accordingly. Oscillations often occur accidentally as a result of damping effects or excitations. In this case, a probability variation of the pump occurs due to the amount of delivery or the hysteresis effect, which makes both the use of the pump difficult for adjustment purposes. Therefore, for the damping of the special valve, the use of a damper in the throttle valve is proposed. In a simple piston damper of the known type, the damping force also generates a negative pressure, which can also cause a cavity which is detrimental to the damping function. This can be eliminated with higher valve pressures when the damper is used. If the damping piston diameter is kept large relative to, for example, the size of the valve diameter, the negative pressure and the additional valve pressure required are reduced. This is desirable because the allowable pressure of the pump is always kept as low as possible.

The possibility of disposing the adjusting element somewhat away from the throttle valve or the individual exhaust chamber makes it possible to combine many or all of the operating elements into a single switching valve with only one drive, It makes the operation alternately possible. In the case of an electric pump adjustment, there is a great advantage in terms of cost and configuration space, since only one adjustment device for many or all of the exhaust chambers is required. In addition, since the amount of fluid contained between the adjusting element and the throttle valve in the channel is substantially non-elastic due to the absence of the cavity, It should not flow in or out to achieve a steady state. Thus, the geometric channel volumes are allowed to vary significantly from one another, which is why the present invention is suitable for all geometric exhaust piston arrangements (for example, axial flow, radial and serial in the case of a piston pump) The location for the adjustment device 27 to be made can be found against all of these.

7 shows several specific features of the throttle valve for the throttle operating element, for example the valve 30 in the first drawing or the valve 150 in the variant according to FIGS. 4 to 6.

However, for the exemplary embodiment of Figures 1, 3, 4 and 5, respectively, having throttle points for each exhaust chamber, this entails high costs.

In the arrangement according to the invention, the pressure difference in the regulating element affects the metered fluid quantity on the basis of the pressure difference. However, at a fixed feed pressure, this pressure difference is reduced to an increasing throttle opening. The use of the pressure difference valve 40 in FIG. 2 shows how this pressure difference can be kept essentially constant, in which the use of a pressure differential valve allows the permissible pressure to flow in parallel with the upstream pressure of the throttle valve Can be changed.

However, the same object can be achieved, at least essentially, in that the spring loaded 2/2-way throttle valve has a sudden opening characteristic, which can be achieved with a soft spring or a large pressure load valve surface or with two coupling, Since p2 is not high enough, the maximum volume flow rate of the pump, i.e. the pressure difference for large valve opening, is not significantly reduced by the regulating valve. Therefore, this measurement basically ensures that the amount of perfusion in the throttle element is only slightly affected by the variance of the spring stiffness or the difference in spring pre-tension or effective valve surface of the inlet valve spring. Thus, this design also eliminates the need for accurate spring sorting or setting of spring pretensioning at each individual inlet valve.

Figures 8, 9 and 10 show the dynamic process of different exhaust chamber charging and operating cycles at the same rotational speed for deformation according to Figures 3 and 4 and 5 respectively, Fig. 2 schematically shows a method according to the present invention. 8 shows the state for full filling or delivery of the exhaust chamber 15 with respect to the upper dead center (TDC) and the lower dead center (BDC) of each exhaust piston 9, 9 shows a state for semi-filling or transferring of the exhaust chamber 15, and FIG. 10 shows a state of delivery of the exhaust chamber 15 other than "0". In FIG. 8 and FIG. 9, the valve end surface inclination (A _valve) and the suction pressure p5 of testosterone in the cylinder before the exhaust chamber charged for the lock (V = V _max) and half-filled gas discharge chamber (V = 0.5V _max) And the pressure difference during charging (P _feed- P _{channel suction} = p2-p3) is stable and is almost the same for all deliveries.

In the present invention, the reason for which may be formed to be immediate and stable way, the 2/2-stroke and the pressure p ₄ of the direction the throttle valve 21 is substantially channel (31a, 31b, 31c or 31a, 31b, 31c or 130a, 130b, 130c, 130d) or cavitation prevented in the regulating device (27 or 150).

As already described in detail above, for a pressure p ₅ in the exhaust chamber 15, in each case, a low value constant in the example close to zero through the bottom dead center (BDC) when the valve is closed .

In this way, during the entire suction operation (valve opening and closing), a constant boundary condition is formed in the form of a substantially constant value for the feed pressure p ₂ and the exhaust chamber pressure p5.

The pipe considering the absence or deficiency of the cavity is achieved by the invention in the (41a, 41b, 41c or 130a, 130b, 130c, 130d) , non-compressible can be assumed between the inlet and exhaust chamber p ₅ with a pressure p ₂ have. Accordingly, the upstream pressure p ₃ of the throttle valve 21 is a result of the continuous condition that perfusion (V ₃₀₎ at the individual throttle cross-section (30) be identical to the perfusion (V ₂₁₎ of the throttle valve 21, a significant delay Lt; / RTI &gt;

here,

α ₂₁ , α ₃₀ , c ₁ , c ₂ are constants defining A ₂₁ (p ₃ )

p is the fluid density.

Therefore, in the inhalation process, the specific A ₂₁ (p ₃ ) and the non-p ₃ are fixedly assigned to the non-A ₃₀ via c ₁ and c ₂ .

Compliance with a fixed assignment may be achieved, for example, by the inherent damper of the arrangement according to the invention already described above as a safeguard against overshooting of the valve 21 during opening and the additional damper 70 described in figure 6A Valve damper.

The pressure p ₃ for various charging situations is close to that of the result p ₂ and p ₅ of the special valve design with reference to FIG. 7. The above equation can be simplified as follows.

A ₂₁ = (? ₃₀ /? ₂₁ ) A ₃₀

In the full filling of the exhaust chamber 15 according to FIG. 8, the free-flowing section A _valve through the inlet valve 28 is assumed to be at its maximum. In half-filling of the exhaust chamber 15 according to FIG. 9, the valve 28 is only partially opened. The transfer volume V corresponds to the area under the volume-flow function. 10 shows a state in which the transfer is exactly zero and the charge becomes zero as it is. In the FIG. 10, as a limiting state, the fluid is still p ₅ = p ₆ at the upper dead center, i.e. it is compressed to the high pressure of the system and released again, but nothing is released in the meantime. Some charge of the exhaust piston occurs to cover any piston leakage as a result of compression / release, even if the delivery volume is " 0 &quot;. Therefore, the minimum opening A _suction (V? 0) is shown in FIG. The duration of this opening extends over almost the entire turn and is only stopped by a relatively short compression / decompression process. Similar trends occur for the other embodiments according to FIGS. 13, 14, 15, 16 and 17, as well as FIGS. 2, 3 and 4.

Fig. 11 shows the transfer-flow characteristic [lacuna] flow rate V = dV / dt as a function of the throttle end face (A _{throttle 1} to A _{throttle 4} ) of the regulating throttle 30 of this control device. The characteristic increases asymptotically from the respective limit rotation speed (C _{m limit 1} ) to (C _{m limit 4} ) and doubles the limit rotation speed for the flow rate. This is because the suction time in addition to the suction cross section also affects, and as is also evident in Figs. 8, 9 and 10, this results in an almost full rotation from the original half- That is, twice.

FIG. 12 shows the corresponding transfer-flow characteristics for a slot-controlled pump as a variation according to FIG. 2.

FIG. 13 shows an embodiment similar to that of FIG. 3, but the design of the 2/2-way throttle valve operated with pressure differential is different and the manner of operation of the regulating valve is different. The two-way throttle valve of the embodiment according to FIG. 13 comprises a ball 54 which is pressed on the valve seat by a spring 53, respectively. Movement in the open position of the valve ball 54 on the valve seat is as a result the filling of the gas discharge chamber is controlled according to the pressure p ₃ depends on the pressure to be distributed to each of the lines (31,31a and 31b). The adjusting device 27 for operating the adjusting valve according to FIG. 1 is suitable for coupling with an analog loop, while the switching valve 50 has an adjusting device according to FIG. 13, which has advantages in relation to digital electronic devices.

The embodiment of FIG. 13 shows this arrangement and an opening angle suitable for the number of slot-controlled pumps and cylinders as in FIG. 2, and the switching valve 50 suffices for a large number of exhaust elements 9.

14 shows an embodiment in which only one 2/2-way throttle valve 81 is used in three exhaust chambers in this example, the 2/2-way throttle valve 81 is installed outside the pump, And is supplied to the individual exhaust chamber 15 through the portions 36, 36a, and 36b.

In this example, the regulating device includes an adjustable exhausting machine 84 that serves as a perfusion volume-limiting function. However, the exhausting machine 84 is preferably driven by an electric machine of variable rotation speed. The venting machine is designed with a constant evacuation machine and obtains a fluid that is delivered directly in line 33 or indirectly through the system 7 to the fluid tank. In this case, the pressure-relief valve has the function of a safety valve or a squirting valve. This prevents an unacceptable increase in the pressure difference in the transfer pressure source if the pressure source is adjusted to a position that delivers more than the maximum absorption of the main pump.

In this example, the 2/2-way throttle valve is spring-loading, not expressed in non-return valve 81, a second-degree valve 21 that is similar to slide valves with no deliberate response of the pressure p in the valve opening ₄ , Then the particular valve opening will remain intact even if it is interrupted during the intake process of the individual exhaust element. In this interruption process, pressure spikes, and p4 is at a pressure p _3, as a result, the ratio of the channel in the cavity (7, 7a, 7b) is reduced in such cases. This interruption process is achieved by selecting the height of the inlet slot 35 in such a way that the inlet slot 35 is opened by the piston 9 below the number of 360 ° per exhaust chamber in each case.

FIG. 15 is very similar to the embodiment of FIG. 14 and likewise uses an adjustable transfer pressure source, here in the form of an adjustable exhaust machine 86, which is driven at a pump rotational speed or a rotational speed proportional thereto. In principle, for example, the driving of the exhaust machine 84 can be achieved by the drive shaft 12 of the pump.

16 shows an embodiment similar to the embodiment of FIG. 3, in which the transfer pressure source 34 returns to a constant speed, wherein control of the inlet pressure adjusts the spring pre- tension of the pressure- That is, the variable pressure-release valve constitutes the adjusting device.

FIG. 17 shows a solution that allows for the separation of the transfer fluid from the tank and the working medium (but the same fluid is also possible).

This form is particularly advantageous when the delivered fluid contains very high viscosity or impurities which can impair its function (e.g., common rail injection device for heavy oil organs), or when the regulating pump is self- It has an advantage when it is operated only at the time. At this time, it is only necessary to have a substantially low powered pressure source 100 for the working fluid, and such pressure sources are already frequently used (e.g., a compressed-air network).

The working fluid reaches the individual 2/2-way throttle valve 103 via line 101 at a controllable pressure p10. In this case, the pressure p10 acts on one working surface 102 of the slide of the 2/2-way throttle valve 103 while the outlet pressure of the spring 104 and the 2/2-way throttle valve is on the line 106 To act on the other working surface 105 of the slide 102.

18 shows a schematic view of a radial pump with three exhaust pistons 9, only the central part of the pump housing is shown around the drive shaft 12 and only the upper exhaust piston 9 is shown completely.

As will be evident, all three exhaust pistons 9 are operated with the common eccentric cam 11, which rotates together with the drive shaft 12.

As is evident from the expression of the upper exhaust piston, this is always in contact with the eccentric cam 11 by the spring 200 like two other exhaust pistons. In the figure, although all three exhaust pistons are driven by the common eccentric cam 11, it is also possible to offset the exhaust piston in the direction of the drive shaft and to drive the exhaust piston with a separate eccentric cam. Other numbers of exhaust pistons may also be selected.

An essential feature of the charge-unregulated pump of FIG. 18 is that the replacement fluid is passed through the inner chamber 202 of the pump housing to the individual exhaust piston 9.

Up to now, the connection line leading to the fluid tank has the reference character 33. Reference numeral 30 denotes an adjustment throttle element which is guided to the inner chamber 202 via a line 31. [ The pump of FIG. 18 is slot-controlled and, for this purpose, has an inlet slot 35 (only the upper exhaust piston is shown), and the inlet slot 35 has a pressure difference (as shown in FIG. 13) Way throttle valve 51 which is operated, and in each case is connected to the corresponding internal combustion chamber 202 and corresponding line parts 204 and 206 in the pump housing. Up to this point, the reference symbol 17 represents an outlet valve which leads through line 18 to the corresponding line of another exhaust piston 9 (not shown) and leads to a " common rail " do.

An opening 208 connected to the inlet slot 35 and connected to the inlet slot at the desired angular range is provided in the exhaust piston 9 to ensure slot control over the corresponding rotation- do.

When the pump is in operation, the individual exhaust pistons 9 are moved up and down in the respective cylinders 210 by the eccentric cams 11 with corresponding springs 200. This allows the fuel to flow through line 33, throttle 30, line 31, inner chamber 202, line 206, 2/2-way throttle valve 51, line 204, 35 and is sucked into the exhaust chamber through the opening portion 208 of the exhaust piston 9 and flows out through the outlet valve 17 under the action of the exhaust piston 9. [

Due to the relatively large volume of the inner chamber 202, it is possible with the present invention to limit the leakage of gas in this interior to such an extent that the pump is fully functional. In this embodiment, it is also possible (not shown) to insert a 2/2-way throttle valve into each exhaust piston 9.

references

(1) Welschof, B.

Analysis of the possibility of application of power control-throttle hydraulic pump by example of fluid hydrostatic second drive in vehicle

(2) Schneider, W.

Pumps for future diesel injection systems

Oil hydraulics and Pneumatics 36 (1992) 5, Pages 304-310

(3) Coper Bessemer Fuel Injection Manual

(4) Schneider, W., Stoclin, M., Lutz T., Eberle, M.

High pressure injection and exhaust gas recirculation

MTZ 54 (1993) 11

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

Gas generation and cavitation in fluid

Journal of Applied Physics 21 (1950) December, pages 1218-1224

(6) Fassbender, A.

Suction throttling - influence of pressure medium and temperature 0

+ P (Oil Hydraulic and Pneumatic) 37 (1993) 9

(7) Schmitt, Th.

Study of Steady and Unsteady Flow through Throttle Section in Fuel Injector of Diesel Engine

Place: TU Munich, 1996

(8) Control and coordination in heating and ventilation engineering

Housing, 1986 Federal Office of Economics 3003 Berue, Swiss

Claims (18)

At least one exhaust chamber for receiving the fluid 2 delivered from the fluid tank 1 having an exposure surface susceptible to the gas pressure p1, preferably atmospheric pressure, (30; 60; 150; 50; 34) arranged to limit the suction of the fluid into the exhaust chamber (15; 129) At least one pressure differential operated two / two-way throttle valve (21; 134; 51; 103) is arranged upstream of the exhaust chamber (15; 129) and downstream of the regulating means (30; 60; 150; 50; , Which is connected to a connecting line (31; 41; 130; 107; 31; 202; 34) between the adjusting means (30; 60; 150; 50; 34) and the 2 / 2- way throttle valve (21; 204 so that the vapor or dissolved gas does not escape from the fluid at least at an absolute pressure of 0.9 bar. The method according to claim 1, There is a pressure source 8, 34 for supplying a fluid at a sufficient high pressure p2 upstream of the adjusting means 30, 60, 150, 50, 34. This means that the fluid from the fluid tank 1 can be directly or indirectly As a control signal. The method according to claim 1, Characterized in that, for an exhaust pump having a plurality of exhaust chambers (15; 129), at least one 2/2-way throttle valve (21; 134; 51; 103) is arranged upstream of the exhaust chamber Lt; / RTI &gt; The method according to claim 1, At least one exhaust chamber (15; 129) comprises at least one inlet slot (35) or at least one inlet valve (28), and at least one 2/2-way throttle valve (21, 21a, , 52, 53, 54; 103) are located very close to the inlet slot (35) or the inlet valve (28). The method of claim 3, And a 2/2-way throttle valve (21; 134; 51; 103) is arranged upstream of each exhaust chamber (15; 129). The method according to claim 1, Characterized in that the adjusting means comprises one of an electric, mechanical, hydraulic and pneumatic regulating throttle valve (30; 150). The method according to claim 1, Characterized in that the adjusting means comprises a combination of a throttle valve (30) and a pressure differential valve (40). The method according to claim 1, Characterized in that the adjusting means comprises an electrically operable pulse width modulated 2/2-way switching valve (60). The method according to claim 1, At least one 2/2-way throttle valve 103 includes its spring-load valve, one operating surface 102 is connected to a pressure of fluid p10 from a pressure source 100 belonging to the second fluid circuit Wherein the control unit is dependent on the control unit. The method according to claim 1, Characterized in that at least one 2/2-way throttle valve is constituted by an inlet valve (134; 51, 52, 53, 54) for spring-loaded and for at least one exhaust chamber (15; 129) . The method according to claim 1, And the damper (70) is engaged with at least one 2/2-way throttle valve (21, 21a, 21b). The method according to claim 3 or 5, Wherein the adjusting means comprises a plurality of regulating throttle valves (30, 30a, 30b; 60; 150), which are operated simultaneously or in groups. The method according to claim 1, The at least one 2/2-way valve (21, 21a, 21b; 134; 51, 52, 53, 54; 103) has a sudden opening characteristic and is supplied to the regulating means (30; 60; 150; (P2) of the fluid is sufficiently large such that the pressure differential over the adjustment means (30; 60; 150; 50; 34) for maximum pump volume flow is substantially unchanged. 13. The method of claim 12, The rotary slide valve 150 includes a regulating throttle valve, which has a housing 138 in which a plurality of chambers are installed, in which there is a hollow rotation or axially displaceable valve member and the holes 155a-155d; 130a-130d ) Are installed for each chamber and are arranged in pairs and are positioned facing each other in the valve member (138). 15. The method of claim 14, Characterized in that the holes (155a-155d; 130a, 130b, 130c, 130d) arranged facing towards the valve member and the housing (138) are made by wire erosion. An exhaust pump with a housing includes an inner chamber (202), at least one exhaust piston (9), and an eccentric cam (11) for operating the exhaust piston (9) A control device according to claim 1, for connecting to at least one exhaust chamber through an appropriate hole in the exhaust passage (35) Characterized in that at least one 2/2-way throttle valve (51) is arranged in the pump housing or at least one exhaust piston (9) upstream of the inlet slot (35). The exhaust pump comprises a plurality of exhaust chambers and a corresponding number of exhaust pistons (9) whereby each exhaust chamber is connected to the control chamber according to claim 16 for connection with the inner chamber (202) In the apparatus, Wherein the adjusting means (30) are common to all the exhaust chambers and are arranged in the lines (33; 31) between the fluid tank (1) and the inner chamber (202). The control device according to claim 1, wherein the exhaust pump comprises a plurality of exhaust chambers and a corresponding number of exhaust pistons (9), each exhaust chamber having an inlet slot (35) for the fluid to be delivered, The inlet slot 35 and the exhaust piston 9 connected thereto are arranged such that the open phase of the inlet slot 35 is divided by the number of the exhaust pistons 9 to be approximately 360 and preferably slightly smaller, 30) to restrict the suction.