EP2422072B1 - High-pressure pump - Google Patents

Description

State of the art

The invention relates to a high-pressure pump, in particular a radial or linear piston pump. Specifically, the invention relates to the field of fuel pumps for fuel injection systems of air-compression, self-igniting internal combustion engines.

From the DE 10 2005 046 670 A1 a high pressure pump for a fuel injector of an internal combustion engine is known. The known high-pressure pump has a multipart pump housing, in which at least one pump element is arranged. The pump element comprises a driven by a drive shaft in a pumping motion pump piston which is slidably guided in a cylinder bore of a portion of the pump housing and defines a pump working space therein. The pump piston is supported on the drive shaft via a hollow-cylindrical ram, wherein the ram is displaceably guided in a bore of a part of the pump housing in the direction of the longitudinal axis of the pump piston. During the suction stroke of the pump piston, in which this moves radially inward, the pump chamber is filled with fuel through a fuel inlet channel with the inlet valve open, with the outlet valve is closed. During the delivery stroke of the pump piston, in which this moves radially outward, fuel is conveyed by the pump piston under high pressure through a fuel discharge channel with the exhaust valve open to a high-pressure accumulator, wherein the inlet valve is closed.

The from the DE 10 2005 046 670 A1 known high pressure pump has the disadvantage that it comes in operation due to the high pressure in the pump working space to a leak on the leadership of the pump piston. This leads to deformations the pump piston and a cylinder bore, in which the pump piston is guided, whereby it comes to a widening of the guide and thus an increasing leakage with increasing high pressure.

Another Hechdruckpumpe for fuel injection systems is from the later published document DE 102008041176A known.

Disclosure of the invention.

The high pressure pump according to the invention with the features of claim 1 has the advantage that a high efficiency is possible. Specifically, the increase in leakage of fuel from the pump working space can be prevented or at least reduced.

The measures listed in the dependent claims advantageous refinements of the claim 1 high-pressure pump are possible.

It is advantageous that the material of the pump piston, which has an anisotropic modulus of elasticity, has a transverse strain number which is not less than 0.3. As a result, a negative radial deformation, in which the pump piston is radially compressed due to the pressure load, can be avoided. This can prevent excessive leakage. A leakage between the pump piston and the cylinder bore, in which the pump piston is guided, has the disadvantage that it comes to efficiency losses, which increase with increasing pressure in the pump chamber. As a result, the scope is limited in terms of relatively high pressures. By reducing the leakage on the one hand the efficiency can be increased. On the other hand, the range of application can be increased towards larger pressures that can be generated by the high-pressure pump.

It is advantageous that the material of the pump piston along the axis of the cylinder bore has a smaller modulus of elasticity than perpendicular to the axis of the cylinder bore. As a result, with increasing pressure in the pump working space, the pump piston is shortened along the axis of the cylinder bore and becomes radially expanded, that is to say positive radial deformation, of the pump piston in relation to a front-end loading of the pump piston. Further, radial impingement of the pump piston by the high pressure fuel occurs in a gap between the cylinder bore and the pump piston. This action counteracts the positive radial deformation of the pump piston. Depending on the configuration of the pump piston, the effectively resulting radial deformation preferably be about zero or greater than zero. As a result, any widening of the cylinder bore occurring in the area of the pump piston can also be compensated.

It is advantageous that the material of the pump piston is a metallic or partially metallic material which is processed anisotropically. In this case, it is further advantageous that the material of the pump piston is processed by at least one anisotropic rolling process and / or at least one anisotropic solidification process. Thus, targeted an anisotropic design of the material in terms of the axial extent of the pump piston and a radial expansion of the pump piston, which is perpendicular to the axial extent of the pump piston done.

It is also advantageous that the material of the pump piston is a glass and / or carbon fiber material which is anisotropically reinforced by glass and / or carbon fibers. As a result, a targeted directional dependence of the modulus of elasticity in the axial direction and in a direction perpendicular to the axial direction can be predetermined in an advantageous manner. In addition, the size of the modulus of elasticity can be specified in the axial direction and in the radial direction.

It is advantageous that the pump piston has an end face, that the end face of the pump piston in the cylinder bore limits the pump working space, and that the pump piston is configured such that when the end face of the pump piston is subjected to a high pressure prevailing in the pump work chamber, the pressure is at least substantially zero Radial deformation of the pump piston at least in the portion of the material having the anisotropic elastic modulus occurs. In this case, it is also advantageous that the material of the pump piston, which has an anisotropic elastic modulus, has a transverse strain number from a range of about 0.3 to about 0.6. In this way, an at least substantially vanishing radial deformation of the pump piston can be achieved in an advantageous manner, so that an increase in leakage with increasing pressure in the pump working space is reduced.

However, it is also advantageous that the pump piston has an end face, that the end face of the pump piston in the cylinder bore limits the pump working space and that the pump piston is designed such that when the end face of the pump piston is acted upon by a high pressure in the pump working space prevailing pressure, a positive radial deformation of the pump piston at least in the portion of the material having the anisotropic elastic modulus occurs. In this case, it is also advantageous that the material of the pump piston, which has an anisotropic elastic modulus, has a transverse strain number that is greater than 0.5. As a result, a positive radial deformation can be achieved in an advantageous manner, so that the increase in leakage with increasing pressure is further reduced or completely prevented. In this case, if necessary, an expansion of the cylinder bore in the region of the pump piston can be compensated in whole or in part by a positive radial expansion.

• Fig. 1 eine Hochdruckpumpe in einer schematischen, axialen Schnittdarstellung entsprechend einem Ausführungsbeispiel der Erfindung;
• Fig. 2 den in Fig. 1 mit II bezeichneten Ausschnitt der Hochdruckpumpe des Ausführungsbeispiels der Erfindung in einer schematischen Darstellung, die eine Längsbelastung veranschaulicht;
• Fig. 3 den in Fig. 2 dargestellten Ausschnitt in einer schematischen Darstellung, die eine Querbelastung veranschaulicht; und
• Fig. 4 den in Fig. 2 dargestellten Ausschnitt in einer schematischen Darstellung, die eine Summe aus der in Fig. 2 veranschaulichten Längsbelastung und der in Fig. 3 veranschaulichten Querbelastung darstellt.

Preferred embodiments of the invention are explained in more detail in the following description with reference to the accompanying drawings, in which corresponding elements are provided with corresponding reference numerals. It shows:

• Fig. 1 a high pressure pump in a schematic, axial sectional view according to an embodiment of the invention;
• Fig. 2 the in Fig. 1 labeled II section of the high-pressure pump of the embodiment of the invention in a schematic representation illustrating a longitudinal load;
• Fig. 3 the in Fig. 2 illustrated section in a schematic representation illustrating a lateral load; and
• Fig. 4 the in Fig. 2 shown detail in a schematic representation, which is a sum of the in Fig. 2 illustrated longitudinal load and the in Fig. 3 illustrates cross-load illustrated.

Fig. 1 shows a high pressure pump 1 in a schematic, axial sectional view according to an embodiment. The high-pressure pump 1 can serve in particular as a radial or series piston pump for fuel injection systems of air-compressing, self-igniting internal combustion engines. Specifically, the high pressure pump 1 is suitable for a fuel injection system having a fuel rail that stores diesel fuel under high pressure. However, the high-pressure pump 1 according to the invention is also suitable for other applications.

The high-pressure pump 1 has a multi-part housing 2. In this embodiment, the housing 2 consists of the housing parts 3, 4, 5, wherein the housing part 3 is a base body 3, the housing part 4 is a cylinder head 4 and the housing part 5 is a fixed to the base body 3 flange 5.

The high pressure pump 1 has a drive shaft 6, which is mounted on the one hand at a bearing point 7 in the housing part 2 and on the other hand at a bearing point 8 in the housing part 3. Between the bearings 7, 8, the drive shaft 6 has a cam 9. The cam 9 may be configured as a single or multiple cam. Further, the term of the cam also includes a configuration of the cam 9 in which the drive shaft 6 has an eccentric portion or the like.

The housing part 3 of the high-pressure pump 1 has a guide bore 12, in which a pump assembly 13 is arranged. The cam 9 is associated with the pump assembly 13. Depending on the configuration of the high-pressure pump 1, it is also possible to provide a plurality of pump assemblies, which are designed in accordance with the pump assembly 13. Such pump assemblies may also be associated with the cam 9 and / or associated with one or more further cams corresponding to the cam 9. Depending on the configuration, a radial or in-line piston pump can thereby be realized.

The cylinder head 4 has a projection 14. The projection 14 extends into the guide bore 12. The projection 14 has a cylinder bore 15 in which a pump piston 16 along an axis 17 of the cylinder bore 15 is slidably guided. The pump piston 16 can be moved back and forth along the axis 17 in the cylinder bore 15, as illustrated by the double arrow 18. The piston 16 has an end face 19 which limits a pump working chamber 20 in the cylinder bore 15. In the pump working chamber 20 performs a fuel passage 21, in which an inlet valve 22 is arranged. During a suction stroke of the pump piston 16, fuel flows from the fuel channel 21 into the pump working chamber 20 via the inlet valve 22. In addition, a fuel channel 23 is provided, in which an outlet valve 24 is arranged. During a delivery stroke of the pump piston 16, fuel under high pressure is conveyed out of the pump working chamber 20 via the outlet valve 24 into the fuel channel 23. The fuel channel 23 is connected, for example, with a fuel rail. As a result, high-pressure fuel can be supplied to such a fuel distributor strip.

The pump assembly 13 has a roller 25 which is mounted in a roller shoe 26. The roller shoe 26 is inserted into a substantially hollow cylindrical plunger body 27. Furthermore, the plunger body 27 is connected to a disc-shaped driving element 28, which surrounds the pump piston 16 above a collar 29. As a result, the pump piston 16 is held in contact with the roller shoe 26 via its collar 29. Here, a plunger spring 30 is provided, which acts on the plunger body 27 and the driving element 28 and thus acts on the plunger body 27 together with the pump piston 16 in the direction of the roller 25 with a certain spring force. As a result, the pump piston 16 with its collar 29, the roller shoe 26, the roller 25 and a running surface 31 of the cam 9 are in each case against each other, this mutual contact is ensured even at high speeds of the high-pressure pump 1.

In operation of the high-pressure pump 1, a reciprocating motion of the pump piston 16 is achieved by a rotation of the drive shaft 6 with the cam 9 about a rotation axis 32 of the drive shaft 6, so that the promotion of fuel under high pressure to a fuel rail or the like on the Fuel channel 23 takes place. During the delivery stroke is thus located in the pump working chamber 20 under high pressure fuel.

The generation of the high pressure in the pump working chamber 20 is thus carried out during operation by means of the moving pump piston 16 which is guided very closely in the cylinder bore 15. Between the cylinder bore 15 and an outer side 35 of the pump piston 16 there is a certain guide gap. The guide gap between the cylinder jacket-shaped outer side 35 of the pump piston 16 and the cylinder bore 15 is chosen so that on the one hand sufficient ease of the pump piston 16 is ensured and on the other hand, a leakage over the guide gap is as small as possible. By this given very narrow leadership low pressure leakage is specified. However, there is the problem that due to the widely varying pressure in the pump working space 20, a pressure-induced widening of the guide gap starting from the pump working space 20 along the cylinder bore 15 may occur. In this case, on the one hand can be expanded in the region of the cylinder bore 15 by acting on the guide gap of the projection 14 of the cylinder head 4. On the other hand, the pump piston 16 can be acted upon on its outer side 35 and compressed accordingly something. The pump piston 16 of the high pressure pump 1 of the embodiment is formed of a material having an anisotropic elastic modulus. As a result, leakage over the guide gap can be reduced. Specifically, a possible increase in the amount of leakage with increasing pressure in the pump working chamber 20 can be prevented or at least reduced. As a result, especially at very high pressures in the pump working chamber 20, an excessively large leakage can be prevented. By occurring at very high pressures in the pump chamber 20 leakage creates a significant loss of energy, which can thus be reduced or prevented. As a result, a high efficiency of the high pressure pump 1 can be ensured even at high pressures to be generated.

The configuration of the pump piston 16 of the high pressure pump 1 of the embodiment is described below with reference to FIGS Fig. 2 to 4 described in further detail.

Fig. 2 shows the in Fig. 1 labeled II section of the pump piston 16 of the high pressure pump 1 of the embodiment in a schematic representation, wherein a radial deformation U _RL is illustrated due to a longitudinal load. In the pump working chamber 20 there is a pressure P, which acts on the end face 19 of the pump piston 16. Along the axis 17 of the pump piston 16 thereby occurs a shortening of the pump piston 16, that is, a change in length 36, on. The pump piston 16 is cylindrical in this embodiment. The pump piston 16 has in this case in the initial state, that is, without pressurization on the end face 19, a radius R. The pump piston 16 is formed of a material having an anisotropic modulus of elasticity. Here, a relatively small elasticity is predetermined along the axis 17, while in the radial direction, that is perpendicular to the axis 17, a relatively large elasticity is predetermined. The modulus of elasticity is therefore relatively small along the axis 17 and relatively large in the radial direction. In addition, the material of the pump piston 16 has a transverse strain number v, which is not smaller than 0.3. The ratio of a longitudinal strain to a transverse strain is thus greater than or equal to 0.3. For example, an anisotropically machined steel may have a transverse strain of 0.3.

The radial deformation U _RL due to the longitudinal load results from the product of the transverse strain number v, the radius R of the pump piston 16 and the quotient of the pressure P and the elastic modulus in the longitudinal direction E _L : $${\mathrm{U}}_{\mathrm{RL}}\mathrm{=}\mathrm{v}\mathrm{\cdot }\mathrm{R}\mathrm{\cdot }\mathrm{P}\mathrm{/}{\mathrm{e}}_{\mathrm{L}}\mathrm{,}$$

This radial deformation U _RL due to the longitudinal load causes a certain expansion of the pump piston 16 along its axis 17, as shown in the Fig. 2 is illustrated by a broken line 37. It should be noted that in the Fig. 2 only the radial deformation U _RL is illustrated due to the longitudinal load, which does not occur in isolation in practice. For the radial deformation U _RL due to the longitudinal load is still a radial deformation U _RQ due to a transverse load, which in the following with reference to Fig. 3 is illustrated.

Fig. 3 shows the in Fig. 2 illustrated pump piston 16 in a partial, schematic representation, wherein the radial deformation U _RQ is illustrated due to transverse load of the pump piston 16. In the Fig. 3 the application of the pump piston 16 on its outer side 35 idealized without the action on its end face 19 is shown. On the outer side 35 of the pump piston 16 along the axis 17 is acted upon by the pressure in the guide gap between the pump piston 16 and the cylinder bore 15. In the area of the end face 19, this pressure is equal to the pressure P in the pump working chamber 20. In a direction 38 along the guide gap, the pressure in the guide gap decreases continuously. This is in the Fig. 3 by arrows of different lengths, which point to the outside 35 of the pump piston 16 illustrated. Due to the pressure in the guide gap, the pump piston 16 is radially compressed, so that the radial deformation U _RQ occurs, which in the Fig. 3 is illustrated by an interrupted line 39. For simplicity, it can be assumed in the zeroth approximation that the pressure in the guide gap in a section 40 of the pump piston 16 adjoining the end face 19 is equal to the pressure P in the pump working chamber 20. The radial deformation U _RQ in section 40 is then obtained as the product of a difference with a minuend equal to the transverse strain number v and a subtrahend equal to the number 1, the pressure P in the pump working space 20 and a quotient with a dividend. which is equal to the radius R of the pump piston 16 and a divisor which is equal to the elasticity in the transverse direction E _Q. The radius R is in this case the radius R of the pump piston 16 in the initial state. This results in: $${\mathrm{U}}_{\mathrm{RQ}}\mathrm{=}\left(\mathrm{v}-1\right)\mathrm{\cdot }\mathrm{P}\mathrm{\cdot }\mathrm{R}\mathrm{/}{\mathrm{e}}_{\mathrm{Q}}\mathrm{,}$$

Fig. 4 shows the in Fig. 2 shown section of the pump piston 16 in a schematic representation for explaining the embodiment, wherein the sum of the basis of the Fig. 2 illustrated radial deformation U _RL on the basis of longitudinal load and the basis Fig. 3 illustrated radial deformation U _{RQ is shown} due to lateral load. While in the FIGS. 2 and 3 Thus, the isolated effects are shown, which serve to design the pump piston 16 is in the Fig. 4 illustrates the combined effect occurring in practice. The radial deformation U results from the sum of the radial deformation U _RL due to longitudinal loading and the radial deformation U _RQ due to transverse loading: $$\mathrm{U}\mathrm{=}{\mathrm{U}}_{\mathrm{RL}}\mathrm{+}{\mathrm{U}}_{\mathrm{RQ}}\mathrm{,}$$

In this case, a shortening of the pump piston 16, that is, a change in length 36, occurs. For the radial deformation U due to longitudinal and transverse loading, formulas (1) and (2) thus result in: $$\mathrm{U}\mathrm{=}\mathrm{P}\mathrm{\cdot }\mathrm{R}\mathrm{\cdot }\left(\mathrm{v}\mathrm{/}{\mathrm{e}}_{\mathrm{L}}\mathrm{+}\left(\mathrm{v}\mathrm{-}\mathrm{1}\right)\mathrm{/}{\mathrm{e}}_{\mathrm{Q}}\right)\mathrm{,}$$

By expanding the formula (4) with a quotient whose dividend and divisor are equal to the elasticity in the longitudinal direction E _L and therefore equal to 1, the result is: $$\mathrm{U}\mathrm{=}\mathrm{P}\mathrm{\cdot }\mathrm{R}\mathrm{\cdot }\left(\mathrm{v}\cdot {\mathrm{e}}_{\mathrm{Q}}\mathrm{+}\mathrm{v}\mathrm{\cdot }{\mathrm{e}}_{\mathrm{L}}\mathrm{-}{\mathrm{e}}_{\mathrm{L}}\right)\mathrm{/}\left({\mathrm{e}}_{\mathrm{L}}\cdot {\mathrm{e}}_{\mathrm{Q}}\right)\mathrm{,}$$

From the formula (4) or the formula (5), it is found that the radial deformation U is positive and large when the elasticity in the longitudinal direction E _L becomes small, when the transverse strain number v becomes large and when the elasticity in the transverse direction E _Q becomes large , Thus, by an embodiment of the pump piston 16 with a predetermined variation of the modulus of elasticity, the desired radial deformation U can be specified.

wobei E den Elastizitätsmodul bezeichnet, der in Längs- und Querrichtung gleich groß ist. In diesem Sonderfall wird die Radialverformung U positiv, wenn die Querdehnungszahl v größer als 0,5 ist. Durch einen weiteren Anstieg der Querdehnungszahl V kann die Radialverformung U weiter vergrößert werden.A special case arises when the elastic modulus in the longitudinal direction E _{L is} equal to the modulus of elasticity in the transverse direction E _Q. Then: $$\mathrm{U}\mathrm{=}\mathrm{P}\mathrm{\cdot }\mathrm{R}\mathrm{\cdot }\left(\mathrm{2}\mathrm{\cdot }\mathrm{v}\mathrm{-}\mathrm{1}\right)\mathrm{/}\mathrm{e}\mathrm{.}$$

where E denotes the modulus of elasticity, which is the same in the longitudinal and transverse directions. In this special case, the radial deformation U becomes positive if the transverse strain number v is greater than 0.5. By a further increase in the transverse strain number V, the radial deformation U can be further increased.

Thus, by a specification in which the elastic modulus in the transverse direction E _{Q is} greater than the modulus of elasticity in the longitudinal direction E _L and / or a specification in which the transverse strain number v is greater than 0.5, a positive and large radial deformation U can be achieved. In this case, it is also possible for the properties of the material of the pump piston 16 to be varied along the axis 17 in the direction 38 in order to achieve adaptation to the pressure also varying along the axis 17 in the guide gap.

By a positive radial deformation U, that is by a radial deformation U, which increases with increasing pressure P in the pump working chamber 20, as it results from the formula (4) and (5), thus, a certain widening of the cylinder bore 15 of the projection 14th be compensated in whole or in part with increasing pressure. Thus, an undesirable increase in leakage over the guide gap can be reduced or prevented. Specifically, leakage occurring during operation can be limited to relatively low values even at high pressures in the pump working space 20. As a result, the high-pressure pump 1 of the exemplary embodiment is also suitable for generating very high pressures in the pump working chamber 20.

Based on Fig. 4 illustrates an embodiment in which the radial deformation U is positive. Depending on the particular application, however, it is also possible to design the pump piston 16 such that it adopts or maintains a cylindrical shape at the high pressure P achieved in the pump working space. This is particularly advantageous in embodiments in which the cylinder bore 15 does not expand or only slightly with increasing pressure P in the pump working chamber 20.

The pump piston 16 may be formed, for example, of a metallic material. In this case, targeted rolling and solidification methods can be used to achieve an anisotropic modulus of elasticity in the pump piston 16. Here, specifically, a difference between the modulus of elasticity in the longitudinal direction E _L , that is along the axis 17, and the elastic modulus in the transverse direction E _Q , that is, in the direction of the radius R, given. Furthermore, it is possible to produce the pump piston 16 from fiberglass and / or carbon fiber materials, wherein the glass and / or carbon fibers are arranged anisotropically, in particular with a larger proportion in the radial direction, in the pump piston 16. Due to the relatively small modulus of elasticity in the longitudinal direction E _L , the deformation 36 under pressure along the axis 17 of the pump piston 16 becomes large, the material of the material of the pump piston 16 preferably deviating in the transverse direction, that is to say in the radial direction, due to the large transverse strain number v. while the pump piston 16 is constricted only very slightly due to the pressure load on its outer side 35 because of the high modulus of elasticity in the transverse direction E _Q.

By a variable design of the modulus of elasticity in the longitudinal and transverse direction, a deformation of the pump piston 16 can be specified under pressure load targeted.

The invention is not limited to the described embodiments.

Claims (10)

1. High-pressure pump (1), in particular radial or in-line piston pump for fuel injection systems of air-compressing, auto-ignition internal combustion engines, having at least one pump assembly (13) and having a drive shaft (6), wherein the pump assembly (13) has a pump piston which is guided in a cylinder bore (15) along an axis (17) of the cylinder bore (15) and which can be driven by the drive shaft (6), and wherein the pump piston (16) delimits, in the cylinder bore (15), a pump working chamber (20),
   characterized
   in that the pump piston (16) is formed at least in portions from a material which has an anisotropic modulus of elasticity.
2. High-pressure pump according to Claim 1,
   characterized
   in that that material of the pump piston (16) which has an anisotropic modulus of elasticity has a Poisson's ratio of not less than 0.3.
3. High-pressure pump according to Claim 1 or 2,
   characterized
   in that the material of the pump piston (16) has a modulus of elasticity which is lower along the axis (14) of the cylinder bore (15) than perpendicular to the axis (14) of the cylinder bore (15).
4. High-pressure pump according to one of Claims 1 to 3,
   characterized
   in that the material of the pump piston (16) is an anisotropically processed metallic or partially metallic material.
5. High-pressure pump according to Claim 4,
   characterized
   in that the material of the pump piston (16) is processed by means of at least one anisotropic rolling process and/or at least one anisotropic hardening process.
6. High-pressure pump according to one of Claims 1 to 3,
   characterized
   in that the material of the pump piston (16) is a glass fibre and/or carbon fibre material which is anisotropically reinforced with glass fibres and/or carbon fibres.
7. High-pressure pump according to one of Claims 1 to 6,
   characterized
   in that the pump piston (16) has an end surface (19), in that the end surface (19) of the pump piston (16) delimits, in the cylinder bore (15), the pump working chamber (20), and in that the pump piston (16) is designed such that, when a high pressure prevailing in the pump working chamber (20) acts on the end surface (19) of the pump piston (16), an at least substantially negligible radial deformation of the pump piston (16) occurs at least in that portion of the material which has the anisotropic modulus of elasticity.
8. High-pressure pump according to Claim 7,
   characterized
   in that that material of the pump piston (16) which has an anisotropic modulus of elasticity has a Poisson's ratio in a range from approximately 0.3 to approximately 0.5.
9. High-pressure pump according to one of Claims 1 to 6,
   characterized
   in that the pump piston (16) has an end surface (19), in that the end surface (19) of the pump piston (16) delimits, in the cylinder bore (15), the pump working chamber (20), and in that the pump piston (16) is designed such that, when a high pressure prevailing in the pump working chamber (20) acts on the end surface (19) of the pump piston (16), a positive radial deformation of the pump piston (16) occurs at least in that portion of the material which has the anisotropic modulus of elasticity.
10. High-pressure pump according to Claim 9,
    characterized
    in that that material of the pump piston (16) which has an anisotropic modulus of elasticity has a Poisson's ratio of greater than 0.5.

Images