Classifications

• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
  • F02M63/00—Other fuel-injection apparatus having pertinent characteristics not provided for in groups F02M39/00 - F02M57/00 or F02M67/00; Details, component parts, or accessories of fuel-injection apparatus, not provided for in, or of interest apart from, the apparatus of groups F02M39/00 - F02M61/00 or F02M67/00; Combination of fuel pump with other devices, e.g. lubricating oil pump
  • F02M63/02—Fuel-injection apparatus having several injectors fed by a common pumping element, or having several pumping elements feeding a common injector; Fuel-injection apparatus having provisions for cutting-out pumps, pumping elements, or injectors; Fuel-injection apparatus having provisions for variably interconnecting pumping elements and injectors alternatively
  • F02M63/0225—Fuel-injection apparatus having a common rail feeding several injectors ; Means for varying pressure in common rails; Pumps feeding common rails
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
  • F02M59/00—Pumps specially adapted for fuel-injection and not provided for in groups F02M39/00 -F02M57/00, e.g. rotary cylinder-block type of pumps
  • F02M59/20—Varying fuel delivery in quantity or timing
  • F02M59/34—Varying fuel delivery in quantity or timing by throttling of passages to pumping elements or of overflow passages, e.g. throttling by means of a pressure-controlled sliding valve having liquid stop or abutment
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
  • F02M59/00—Pumps specially adapted for fuel-injection and not provided for in groups F02M39/00 -F02M57/00, e.g. rotary cylinder-block type of pumps
  • F02M59/20—Varying fuel delivery in quantity or timing
  • F02M59/36—Varying fuel delivery in quantity or timing by variably-timed valves controlling fuel passages to pumping elements or overflow passages
  • F02M59/365—Varying fuel delivery in quantity or timing by variably-timed valves controlling fuel passages to pumping elements or overflow passages valves being actuated by the fluid pressure produced in an auxiliary pump, e.g. pumps with differential pistons; Regulated pressure of supply pump actuating a metering valve, e.g. a sleeve surrounding the pump piston
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
  • F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
  • F04B49/22—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves
  • F04B49/225—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves with throttling valves or valves varying the pump inlet opening or the outlet opening
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
  • F02B3/00—Engines characterised by air compression and subsequent fuel addition
  • F02B3/06—Engines characterised by air compression and subsequent fuel addition with compression ignition
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02D—CONTROLLING COMBUSTION ENGINES
  • F02D41/00—Electrical control of supply of combustible mixture or its constituents
  • F02D41/30—Controlling fuel injection
  • F02D41/38—Controlling fuel injection of the high pressure type
  • F02D41/3809—Common rail control systems
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
  • F04B2205/00—Fluid parameters
  • F04B2205/06—Pressure in a (hydraulic) circuit
  • F04B2205/062—Pressure in a (hydraulic) circuit before a throttle
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
  • F04B2205/00—Fluid parameters
  • F04B2205/08—Pressure difference over a throttle

Definitions

• 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.
• 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.).
• variable pumps 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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 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.
• 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.
• 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.
• 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.
• n-cylindrical pump has a high periodicity, i.e.
• 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.
• 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 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 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.
• a pressure p1 for example when resting in a tank ventilated to the atmosphere
• 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 ⁇ 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.
• 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 ⁇ 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.
• 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.
• 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 ⁇ ptemp.
• p x (k (T ⁇ / k (T ⁇ ) -l) p 1 if k (T ⁇ ) ⁇ k (T 1 ) , in which T ⁇ 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.
• the assigned, full response to the delivery rate takes place with the first subsequent, complete suction process (in principle, it could'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.
• FIG. 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
• 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,
• FIG. 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
• FIG. 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,
• FIG. 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,
• FIG. 13 shows an embodiment of a control device according to the invention with a switching valve as an adjusting device
• FIG. 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,
• 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
• FIG. 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
• FIG. 18 shows a schematic view of a further degree of filling variable pump according to the invention.
• FIG. 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.
• 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.
• 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.
• 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.
• 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.
• 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 ⁇ p 3 which counteracts overshoot - because of the lack of voids according to the invention!
• 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.
• 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.
• 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).
• 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 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• FIGS. 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.
• 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.
• a respective inlet valve 134 and a respective outlet valve 118 are provided for each displacement space 129a-d (of which the displacement space 129c is not shown).
• 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.
• each housing part 133a-d 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.
• valve 150 integrated into the construction, which is designed, for example, in accordance with DE-PS 37 14 691 can be.
• 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 .
• each distributor bore 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.
• 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.
• an electric drive 158 FIG.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.
• the need for only one converter for several or all displacement spaces is a great advantage in terms of cost and installation space.
• 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.
• 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.
• FIG. 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.
• the pressure difference at the adjusting element enters the metered amount of liquid with the root of the pressure difference.
• 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.
• 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.
• FIGS. 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.
• incompressibility may be assumed between inflow with pressure p 2 and displacement space p 5 .
• 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 :
• 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 .
• the free flow cross section A valve through the inlet valves 28 assumes the maximum value.
• 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.
• 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. 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.
• 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.
• the adjusting device consists of an adjustable displacement machine 84 which has a flow-limiting function.
• 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.
• 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.
• 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.
• 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.
• 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.
• FIG. 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.
• the latter like the two further displacement pistons, is always kept in contact with the eccentric cam 11 by means of a spring 200.
• 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.
• 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.
• 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.
• 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.
• 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.
• 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.

Landscapes

• Engineering & Computer Science (AREA)
• Mechanical Engineering (AREA)
• General Engineering & Computer Science (AREA)
• Chemical & Material Sciences (AREA)
• Combustion & Propulsion (AREA)
• Fluid Mechanics (AREA)
• Physics & Mathematics (AREA)
• Details Of Reciprocating Pumps (AREA)
• Fuel-Injection Apparatus (AREA)
• Rotary Pumps (AREA)
• Reciprocating Pumps (AREA)
• Steering Control In Accordance With Driving Conditions (AREA)
• Vehicle Body Suspensions (AREA)
• Control Of Non-Positive-Displacement Pumps (AREA)

Applications Claiming Priority (4)

Publications (2)

Family

ID=25678018

Family Applications (1)

Country Status (9)

Cited By (2)

Families Citing this family (62)

Family Cites Families (21)

• 1994
  • 1994-11-07 US US08/464,856 patent/US5701873A/en not_active Expired - Lifetime
  • 1994-11-07 JP JP51350995A patent/JP3747061B2/ja not_active Expired - Lifetime
  • 1994-11-07 AT AT94930902T patent/ATE169720T1/de active
  • 1994-11-07 ES ES94930902T patent/ES2120076T3/es not_active Expired - Lifetime
  • 1994-11-07 CN CN94190910A patent/CN1082143C/zh not_active Expired - Lifetime
  • 1994-11-07 DE DE59406680T patent/DE59406680D1/de not_active Expired - Lifetime
  • 1994-11-07 WO PCT/CH1994/000215 patent/WO1995013474A1/fr not_active Ceased
  • 1994-11-07 EP EP94930902A patent/EP0678166B1/fr not_active Expired - Lifetime
• 1995
  • 1995-06-07 CA CA002151518A patent/CA2151518A1/fr not_active Abandoned

Non-Patent Citations (1)

Also Published As

Similar Documents

Legal Events