EP1655480B1 - Method of using a fuel system of an internal combustion engine and fuel system - Google Patents

Description

State of the art

The invention relates firstly to a method for operating a fuel system of an internal combustion engine in which at least one displacement element of a fuel pump is driven by a drive device and performs a displacement and a suction movement within a delivery cycle, and in which during the displacement movement may be present during the promotion Fuel is conveyed in a high-pressure area.

The subject of the present invention is also a fuel system with a fuel pump with at least one displacement element, which is driven by a cam device.

A method and a fuel system of the type mentioned are known from the market. They are used in fuel systems of internal combustion engines with direct fuel injection. In doing so, the fuel pump feeds into a fuel rail (common rail), in which the fuel is stored under high pressure is. In the known fuel pump is a fuel pump whose displacement element is formed by a piston which is driven by a camshaft.

To set the flow rate, a quantity control valve is provided which can temporarily connect the delivery space to an inlet-side low-pressure region during a delivery movement of the piston. From the DE 100 52 629 A1 and belonging to the patent family US 2002/0053338 A1 It is known that the quantity control valve is closed at the beginning of a conveying movement of the piston, so that is conveyed into the fuel rail and that, depending on the amount of fuel to be delivered, the quantity control valve opens in the course of the conveying movement, so that the fuel is no longer in the Fuel manifold, but back to the low pressure range is promoted. Also known from the market is that the quantity control valve is open at the beginning of the conveying movement, so that initially is not promoted in the fuel rail. In the course of the displacement movement closes the quantity control valve, so that can be promoted from this point in the fuel rail and the high pressure area.

The EP 0 304 741 A1 shows a piston pump with a driven by a drive device with a trajectory displacement element, wherein the trajectory is shaped so that the maximum speed of the displacement element during a delivery phase over a range is constant.

As a general state of the art is further on the EP 1 072 787 A2 , the EP 0 481 964 A2 , the DE 42 23 728 A1 and the GB 1 249 942 A directed.

Object of the present invention is to provide a method and a fuel system of the type mentioned in such a way that a corresponding fuel system can be built inexpensively and compact and easily manufactured.

This object is achieved in a method of the type mentioned above in that the displacement element is driven so that during the displacement movement is a phase at the same time substantially constant and in the displacement direction substantially maximum speed of the displacement element, wherein a trajectory of the drive means in the region bottom dead center of the displacement element a concave. Curvature radius has. In a fuel system of the type mentioned, the stated object is achieved in that the displacement element operates according to the method described above.

Advantages of the invention

Due to the concave radius of curvature of the trajectory of the drive device in the region of the bottom dead center of the displacement element is achieved that quickly a maximum piston speed is reached, or one can decelerate the displacement element within a few degrees and thus over a large angle, the displacement element with relatively high Speed can move. By quickly reaching the maximum velocity of the displacement element and being constant during the displacement movement for a certain period of time, the amount of maximum velocity at a given displacement of the displacement element can be kept comparatively low. A low maximum speed of the Displacement element during the displacement movement, however, has several advantages: on the one hand pressure pulsations are reduced during the displacement movement. The corresponding components downstream and / or upstream of the fuel pump may therefore meet less stringent strength requirements and thus be made more compact and less expensive.

Due to the comparatively low velocity of the displacement element during the displacement movement is obtained at an open valve, through which the fuel flows from the delivery chamber of the fuel pump, a lower flow rate of the fuel and a correspondingly lower flow force acting on the valve. Also, the valve itself can therefore be smaller and easier to build and thus manufactured inexpensively.

Advantageous developments of the invention are specified in subclaims.

In a first development of the method according to the invention, it is proposed that the phase of essentially constant speed occupy at least approximately 50% of the phase of the displacement movement. From this duration of the constant speed phase, the reduction of pulsations upstream and / or downstream of the fuel pump is particularly significant, as well as the reduction of the flow forces acting on a valve delimiting the delivery space.

It is optimal if the phase with constant and maximum speed of the displacement element takes as long as possible. However, the longer this phase lasts, the greater must be the positive acceleration with which the displacement element is brought from the bottom dead center to the said speed. However, a high acceleration of the displacement element leads to high forces between the displacement element and the drive device, with which the displacement element is driven. If a phase in which the displacement element is positively accelerated to the substantially constant speed assumes at most approximately 15% of the phase of the displacement movement, the forces acting between the displacement element and the drive device are still sufficiently low and at the same time sufficiently high Acceleration that a long phase with constant and maximum speed of the displacement element can be achieved.

It is advantageous if the positive acceleration is substantially constant, since in this way the forces between the displacement element and the drive device can be kept comparatively low. A similar effect occurs when the velocity of the displacement element with substantially constant negative acceleration drops from the maximum velocity.

It is also particularly advantageous if the magnitude of a mean positive acceleration, with which the displacement element is accelerated to the substantially constant maximum speed, is greater than the amount of a mean negative acceleration, with which the displacement element is decelerated from the substantially constant speed. This is based on the consideration that in conventional pumps, the displacement element is acted upon by a clamping device against the drive means. This is intended to prevent the displacement element from lifting off the drive element in the region of negative accelerations. The force of inertia acting on the drive element and on the displacement element is the higher, the higher this negative acceleration is. By reducing this negative acceleration, therefore, this inertia force can be reduced. This in turn allows for a reduction of the said tensioning device, which overall benefits the compactness of the fuel pump.

A comparatively small negative acceleration has yet another advantage: a large negative acceleration in the area of top dead center of the displacement element would require a comparatively small radius of curvature in this area in the case of a corresponding drive device, for example a drive cam. This in turn would result in a high surface pressure between the displacement element and the drive device. By reducing the negative acceleration, the surface pressure between the displacement element and the drive device is thus reduced in the region of top dead center, which in turn simplifies the construction of the fuel pump and increases its service life.

It has been found to be particularly favorable if the magnitude of the average positive acceleration is at least a factor of 2 greater than the amount of the average negative acceleration. This results in a sufficiently long phase with a constant maximum speed of the displacement element with at the same time not too high demands on the structural strength, in particular in the region of the interface between the displacement element and the drive device.

It is also particularly favorable when the displacement element is driven by a cam device and the course of a negative acceleration, with which the displacement element is decelerated from the substantially constant speed, is selected so that a resulting therefrom on the rotation angle of the cam device course of a the displacement element acting mass force at least approximately and at least in Area of the top dead center of the displacement element corresponds to the course of a force acting on the displacement element to the cam device towards spring force, optionally taking into account a safety factor. In this area, therefore, the acceleration profile of the displacement element is adapted to the force curve of the spring force acting on the displacement element. Thus, this spring force is optimally utilized over a comparatively wide range. This in turn allows an extension of that phase in which the speed of the displacement element is maximum and constant, which in turn benefits the reduction of the maximum speed.

It is also particularly advantageous if the duration of the displacement movement within a delivery cycle is greater than the duration of the suction movement. Such an asymmetrical design of the lift curve of the displacement element means a shift of the top dead center towards a larger angle of rotation of the drive device. An extension of the duration of the displacement movement allows a further reduction of the maximum speed of the displacement element during the displacement movement. It is proposed in a concrete embodiment that the duration of the displacement movement is at least 15% greater than the duration of the suction movement.

The fact that the angle in the region of the concave radius of curvature at the end of the suction movement is greater than that Angle of the area of the concave radius of curvature at the beginning of the displacement movement, one achieves that one accelerates and decelerates the displacement element with similarly large acceleration forces and thus can optimally exploit the strength limits of the pump.

Particularly advantageous are the advantages of the method according to the invention, when the beginning of the delivery phase depends on a controllable valve device, which can forcibly connect a delivery chamber of the fuel pump with a low pressure region. Would the beginning of the funding phase in an operating phase of the fuel pump fall, in which the displacement element moves at high speed, there would be a particularly strong Pulsationsanregung in the high pressure region of the fuel system. By limiting the speed of the displacement element these Pulsationsanregungen are therefore significantly reduced.

In this case, this embodiment of the method according to the invention has a further advantage: when the delivery chamber of the fuel pump is connected to the low-pressure region via the controllable valve device, the fuel is expelled from the delivery chamber via the valve device into the low-pressure region. Accordingly, a flow force acts on the valve device. In a normally closed valve device, the flow force directly affects the force required to hold open. By reducing the flow force so also this required force is reduced to keep open, which allows the use of a smaller actuator in the valve device. In a normally open valve device, the high flow force requires a high spring force to keep the valve device open and consequently a high actuation force for closing the valve device. Again, therefore, the reduction of the flow force leads to a more compact design of the valve device.

drawings

Figur 1

eine schematische Darstellung eines Kraftstoffsystems einer Brennkraftmaschine mit einer Kraftstoffhochdruckpumpe und einem Mengensteuerventil;

Figur 2

eine schematische Darstellung der Kraftstoffhochdruckpumpe und des Mengensteuerventils von Figur 1 während unterschiedlicher Betriebszustände;

Figur 3

eine vergrößerte Darstellung eines Nockens, mit dem ein Verdrängungselement der Kraftstoffhochdruckpumpe von Figur 2 angetrieben wird;

Figur 4

ein Diagramm, in dem ein Hub, eine Geschwindigkeit, und eine Beschleunigung des Verdrängungselements der Kraftstoffhochdruckpumpe von Figur 1 über dem Drehwinkel des Nockens von Figur 3 aufgetragen sind;

Figur 5

ein Diagramm, in dem eine auf das Verdrängungselement der Kraftstoffhochdruckpumpe von Figur 1 wirkende Massenkraft und einer Federkraft über dem Drehwinkel des Nockens von Figur 3 aufgetragen sind; und

Figur 6

ein Diagramm, in dem ein Krümmungsradius des Nockens von Figur 3 über seinem Drehwinkel aufgetragen ist.

Hereinafter, a particularly preferred embodiment of the present invention will be explained in more detail with reference to the accompanying drawings. In the drawing show:

FIG. 1

a schematic representation of a fuel system of an internal combustion engine with a high-pressure fuel pump and a quantity control valve;

FIG. 2

a schematic representation of the high-pressure fuel pump and the quantity control valve of FIG. 1 during different operating conditions;

FIG. 3

an enlarged view of a cam, with a displacement element of the high-pressure fuel pump of FIG. 2 is driven;

FIG. 4

a diagram in which a stroke, a speed, and an acceleration of the displacement element of the high-pressure fuel pump of FIG. 1 above the rotation angle of the cam of FIG. 3 are applied;

FIG. 5

a diagram in which a on the displacement element of the high-pressure fuel pump of FIG. 1 acting mass force and a spring force over the rotation angle of the cam of FIG. 3 are applied; and

FIG. 6

a diagram in which a radius of curvature of the cam of FIG. 3 is applied over its angle of rotation.

Description of the embodiments

In FIG. 1 a fuel system carries the overall reference numeral 10. It comprises a fuel tank 12 from which an electrically driven feed pump 14 promotes the fuel in a low-pressure line 16. This leads to an inlet 18 of a high-pressure fuel pump 20, which is mechanically driven by an internal combustion engine, not shown in detail. The mechanical connection carries in FIG. 1 an outlet 24 of the high-pressure fuel pump 20 leads to a fuel rail 26, which is also referred to as "common rail" and in which the fuel is stored under high pressure. To the fuel manifold 26 a plurality of injectors 28 are connected, which inject the fuel in them associated combustion chambers 30. The amount of fuel delivered by the high pressure fuel pump 20 to the fuel rail 26 is adjusted inter alia by a valve means 32, referred to as a quantity control valve, as now with reference to FIG FIG. 2 explained in detail:

How out FIG. 2 As can be seen, the high-pressure fuel pump 20 comprises a housing 34, in which a delivery chamber 36 and a displacement element in the form of a piston 38 are provided. This is acted upon by a piston spring 40 via a roller 42 against a cam device 44. This comprises three at an angle of 120 ° to each other arranged cam 46, but only one of which is provided with a reference numeral.

The delivery chamber 36 can be connected to the outlet 24 via a spring-loaded outlet valve 48. The Quantity control valve 32 is also the inlet valve, by which at least temporarily, the low pressure line 16 can be connected to the pumping chamber 36. For this purpose, the quantity control valve 32 has an electromagnetic actuator 50 and a pull-valve spring 52.

In de-energized state, the Mengensteuer- or inlet valve 32 is closed under the action of the tension spring 52. During a suction movement of the piston 38 (reference numeral 54 in FIG FIG. 2 ), the quantity control valve 32 is opened by energizing the electromagnetic actuator 50 against the force of the valve spring 52 and due to the pressure difference between the inlet 18 and the delivery chamber 36 so that fuel from the inlet 18 into the delivery chamber 36 can pass (left illustration of the pump 20 in FIG FIG. 2 ). The exhaust valve 24 is closed during this time.

At the beginning of a displacement movement (reference numeral 56 and middle illustration of the pump 20 in FIG FIG. 2 ), the quantity control valve 32 is further forcibly opened by energizing the electromagnetic actuator 50, so that the fuel is not discharged to the outlet 24, but back to the inlet 18 into the low pressure line 16. During the actual delivery phase, while the fuel is delivered via the outlet valve 48 to the outlet 24 and further into the fuel rail 26, the mass control valve 32 is closed (reference numeral 58 and right hand representation of the pump 20 in FIG FIG. 2 ). A complete suction and displacement movement result in a delivery cycle 59.

As mentioned above, the cam device 44 has a total of three cams 46, the trajectory of which in turn can each be subdivided into four regions A, B, C and D. These areas are in the FIGS. 3 to 6 separated by double dot-dash lines. In FIG. 4 is a the stroke of the piston 38 on the rotational angle of the cam device 44 descriptive curve with the reference numeral 60, a speed of the piston 38 descriptive curve with the reference numeral 62, and a piston 38 describing the acceleration of the curve designated by the reference numeral 64. In FIG. 5 carries a curve which corresponds to a force acting on the piston 38, the mass force, the reference numeral 66, whereas a curve which describes the course of a force acting on the piston 38 by the piston spring 40 spring force, the reference numeral 68 carries. The curve describing the radius of curvature of cam 46 bears in FIG FIG. 6 the reference numeral 70.

One recognizes in particular FIG. 4 in that the area A of the cam path of the cam 46 is such that the piston 38 experiences an overall substantially constant positive acceleration from a bottom dead center UT (curve 64), resulting in an overall substantially linear velocity increase (curve 62) , The acceleration of the piston 38 in the region A of the trajectory of the cam 46 should be as large as possible. It is limited by the criterion of the smallest concave radius of curvature of the trajectory: a large acceleration leads to a small concave curvature radius. The aim in the design of this area A is to choose the course of the acceleration, ie the trajectory of the cam 46 in the area A so that the resulting radius of curvature is as close as possible to the minimum allowable radius of curvature.

Usually, the cam track in this area A has the curvature of a grinding wheel, with which the cam track of the cam 46 is made, or the smallest radius of curvature that can be made with such a grinding wheel. In FIG. 3 Such a grinding wheel is indicated by the reference numeral 72 in dashed lines.

The area B of the trajectory of the cam 46 adjoining the area A is characterized by a constant lifting speed (curve 62 in FIG FIG. 4 ) of the piston 38. This speed is also the maximum positive velocity of the piston 38 during the displacement movement. The amount of the maximum speed in the area B should be as small as possible in order to minimize the flow forces on the inlet or quantity control valve 32 and thus the required magnetic force of the electromagnetic actuator 50. In addition, a speed-dependent pressure excitation in the fuel rail 26 can thereby be kept low.

The area under the curve 62 between bottom dead center UT and top dead center OT corresponds to the stroke of the piston 38. At a given stroke, the maximum speed in the area B can then be as low as possible, if the phase in which the speed is constant ( Area B), takes as long as possible within a delivery cycle 59 of the high-pressure fuel pump 20. In the present case, this is achieved on the one hand by the fact that the positive acceleration (range A) is as large as possible, and thus the duration of the range A is as short as possible, and, on the other hand, by an overall asymmetric design of the Cam 46, through which the top dead center OT is shifted backward in time.

Within the in the FIGS. 4 to 6 Thus, the conveying cycle shown thus extends the displacement movement of the piston 38 from a cam angle of 0 ° to about 65 °, whereas a suction movement (reference numeral 54 in FIG FIG. 2 ) in the present cam 46 lasts from a cam angle of 65 ° to 120 °. This gives you more time during the displacement movement to reach the desired stroke. Because of this longer time, in turn, the maximum speed in area B may be lower. In a cam device, not shown here, which has only two cams, the top dead center OT would be shifted to a cam angle greater than 90 ° and at a cam device having four cams to a cam angle greater than 45 °.

The constant and maximum speed phase B of the piston 38 in the present embodiment extends from a cam angle of about 9 ° to a cam angle of about 44 °. This corresponds to approximately 54% of the total displacement movement. The acceleration phase A corresponds to approximately 14% of the total displacement movement.

In the region C of the trajectory of the cam 46 and the lifting cam 60 of the piston 38, the piston 38 is decelerated, so he experiences a negative acceleration. However, the amount of the mean value thereof in the present embodiment is less than half the magnitude of the average value of the positive acceleration during the phase A. Due to the negative acceleration in the range C, in particular near the top dead center OT, there is a risk that of the Piston 38 lifts off with the roller 42 from the cam 46. This lifting is intended by the piston spring 40 and the corresponding spring force (reference numeral 68 in FIG FIG. 5 ) which prevents the lifting mass inertia force (reference 66 in FIG FIG. 5 ) counteracts. For structural reasons, it is desirable to design the piston spring 40 as small as possible, ie to keep the required spring force as low as possible. This is achieved by the lowest possible negative acceleration in the area C.

An additional advantage of a comparatively small negative acceleration in the region C relates to the surface pressure between the roller 42 and the cam 46. It must be ensured that a permissible value, which also results from the material combination between cam 46 and roller 42, is not exceeded , In addition to the contact forces occurring and the radius of the roller 42, the radius of curvature of the cam 46 is decisive for the height of the surface pressure. A strong negative acceleration in the region of top dead center TDC means a correspondingly small radius of curvature of the trajectory of the cam 46 in this region a correspondingly high surface pressure between roller 42 and cam 46 would lead. For a correspondingly low surface pressure, therefore, the negative acceleration in the region C is also to be kept low. The trajectory of the cam 46 is chosen so that the resulting from the angle of the cam 46 course of the inertia force 66 approximately corresponds to the course of the force acting on the piston 38 spring force 68, taking into account a safety factor of 50% (reference numeral 74 in FIG. 5 ). As a result, the force of the piston spring 40 is optimally utilized over a wide range. This also allows for a certain extension of the design range B and the Reduction of the maximum speed level available there.

For the region D of the trajectory of the cam 46 and the lifting curve 60 of the piston 38, the same design criteria apply as for the region A, ie the highest possible and constant positive acceleration of the piston 38.

Claims (12)

1. Method for operating a fuel system (10) of an internal combustion engine, in which at least one expulsion element (38) of a fuel pump (20) is driven by a drive device (44) via a path curve of the drive device (44) and executes an expulsion movement (56) and an intake movement (54) during one feed cycle, and in which during the expulsion movement (56) there may be a feed phase (58) during which fuel is fed into a highpressure region (26), and wherein the expulsion element (38) is driven in such a way that during the expulsion movement (56) there is a phase (B) with simultaneously essentially constant speed (62) of the expulsion element (38) and essentially maximum speed (62) of the expulsion element (38) in the expulsion direction, characterized in that the path curve of the drive device (44) has a concave curvature radius in the region of the bottom dead centre (UT) of the expulsion element (38), and in that an angle (D) of the region of the concave curvature radius at the end of the intake movement (54) is greater than an angle (A) of the region of the concave curvature radius at the start of the expulsion movement (56).
2. Method according to Claim 1, characterized in that the phase (B) with an essentially constant speed (62) takes up at least approximately 50% of the phase of the expulsion movement (56).
3. Method according to one of Claims 1 or 2, characterized in that a phase (A) in which the expulsion element (38) is positively accelerated to the essentially constant speed (62) takes up at most approximately 15% of the phase of the expulsion movement (56).
4. Method according to Claim 3, characterized in that the positive acceleration (64) is essentially constant.
5. Method according to one of the preceding claims, characterized in that the speed (62) of the expulsion element (38) drops from the maximum speed (62) with an essentially constant negative acceleration (64).
6. Method according to one of the preceding claims, characterized in that the absolute value of an average positive acceleration (64) with which the expulsion element (38) is accelerated to the essentially constant speed (62) is greater than the absolute value of an average negative acceleration (64) with which the expulsion element (38) is braked from the essentially constant speed (62).
7. Method according to Claim 6, characterized in that the absolute value of the average positive acceleration (64) is greater than the absolute value of the average negative acceleration (64) by at least a factor of 2.
8. Method according to one of Claims 6 or 7, characterized in that the expulsion element (38) is driven by a cam device (44) and the profile of a negative acceleration (64) with which the expulsion element (38) is braked from the essentially constant speed (62) is selected such that a profile (66) resulting from over the rotational angle of the cam device (44), of a force of inertia acting on the expulsion element (38) corresponds at least approximately and at least in the region of the top dead centre (OT) of the expulsion element (38) to the profile (68) of a spring force acting on the expulsion element (38) in the direction of the cam device (44), if appropriate taking into account a safety factor (74).
9. Method according to one of the preceding claims, characterized in that the duration of the expulsion movement (56) within one feed cycle is longer than the duration of the intake movement (54).
10. Method according to Claim 9, characterized in that the duration of the expulsion movement (56) is at least 15% longer than the duration of the intake movement (58).
11. Method according to one of the preceding claims, characterized in that the start of the feed phase (58) depends on a controllable valve device (32) which can forcibly connect a feed space (36) of the fuel pump (20) to a low pressure region (16).
12. Fuel system (10), having a fuel pump (20) with at least one expulsion element (38) which is driven by a cam device (44), characterized in that the cam device (44) is shaped in such a way that the expulsion element (38) operates according to one of the methods in claims 1 to 11.

Images