WO2005026528A1

WO2005026528A1 - Hydraulic compensation element

Info

Publication number: WO2005026528A1
Application number: PCT/EP2004/052134
Authority: WO
Inventors: Georg Bachmaier; Bernhard Fischer; Bernhard Gottlieb; Andreas Kappel; Tim Schwebel
Original assignee: Siemens Aktiengesellschaft
Priority date: 2003-09-12
Filing date: 2004-09-10
Publication date: 2005-03-24

Abstract

Disclosed is a closed hydraulic compensation element which can be filled with fluid and which comprises the following characteristics: a hollow-cylindrical housing (2) with an inner chamber which has at least two different cross-axial diameters and the inner wall of the housing has at least one axially pressure-effective surface (2a'); a sealable filling bore (11) leading to the inner chamber of the housing; a piston (3) which is received inside the housing (2) in an axially displaceable manner and provided with at least one pressure-effective surface located opposite the axially pressure-effective surface (2a') of the housing; an outer contour which is adapted to the inner wall of the housing in such a way that loose clearances (6, 7) are partially formed, throttling the fluid, between the piston (3) and the housing (2).

Description

description

Hydraulic compensation element

The invention relates to a hydraulic compensation element.

Mechanical tolerances, temperature-related and pressure-related changes in length, aging effects of a PMA (piezoelectric multilayer actuator) used in particular in a fluid valve, hereinafter referred to as "piezo actuator", have a direct effect on the opening stroke of a fluid valve connected to the piezo actuator and thus on its metered quantity , Conventional methods for compensating the temperature-related change in length of the piezo actuator using suitable material combinations, however, pose serious stability and manufacturing problems.

The elongation ratio of the piezo actuator that can be achieved due to the inverse piezoelectric effect in high-performance ceramics due to the application of a maximum field strength of approx. 2KV / mm permissible for continuous operation is only 1.2-1.4 μm elongation per 1 mm length of the piezo actuator. With a typical length of the piezo actuator of approx. 40mm and a piezo layer spacing of 80μm at 160V applied voltage, the inverse piezoelectric effect leads to an elongation of maximum 56μm. So if there is only a minimal relative deviation in the effective temperature expansion coefficient of approx. 1 * 10 ^{~ 6} 1 / K over the length of the piezo actuator of 40mm between the piezo actuator and the housing in which the piezo actuator is installed, this leads in the automotive-relevant temperature range -40 ° C to 140 ° C to a deviation of the reference areas relevant for valve operation from - 2.4μm to + 4.8μm or in total to 7.2μm and related to the elongation of the piezo actuator to a deviation bandwidth of up to 13%. This creates problems of the following kind: the complex manufacture of the piezo actuator in such narrow tolerances as described above that its thermal expansion remains in a sufficiently narrow tolerance range is practically not feasible as a component with a domain structure and hysteresis, the temperature expansion coefficient depends strongly on the state of polarization and the mechanical and electrical load history of the piezo actuator. The temperature dependence of the length of the piezo actuator is non-linear. The temperature expansion coefficient can assume values in the range from -5 * 10 ^"6 1 / K to + 7 * 10 ^{" 6} 1 / K for the same piezo actuator.

From DE 19940055 Cl a metering device is known which can ensure both a length compensation and a stiff power transmission. The slow length compensation takes place through a long-lasting overpressure in a hydraulic chamber (temperature expansion) which lies between a piston connected to a piezo actuator and a fixed housing part. The overpressure leads to hydraulic compensation via a fit which leads from the hydraulic chamber to an actuator space formed around the piezo actuator.

The present invention is based on the object of specifying a device which is stiff in comparison to short-term application of force with a small space requirement and achieves a length compensation of an element acting on said device.

The object is solved by the features of the independent claims.

A hydraulically closed hydraulic compensation element is proposed which can be filled with a fluid and which has the following further features: a hollow-cylindrical housing with an interior which has at least two different transverse axial diameters and the inner wall of the housing has at least one axially pressure-effective surface, a closable filling bore leading into the interior of the housing and a piston which is axially displaceably received in the housing and has - at least one axially pressure-effective surface which is opposite the axially pressure-effective surface of the housing, an outer contour which is adapted to the inner wall of the housing in such a way that partial clearance fits with a fluid-throttling effect are formed between the piston and the housing, at least one hydraulic chamber which is located between the at least one axially pressure-effective surface of the Piston and the at least one axially pressure-effective surface of the housing and is connected on both sides to the clearance fits, elastic membranes which are attached to the housing such that Speic Hervolumina are formed between the membranes and the housing, wherein at least one elastic membrane connects the housing to the piston on each end face of the housing.

Thanks to its clearance fits and hydraulic chambers, the hydraulic compensation element achieves a high level of rigidity when subjected to force. The elastic diaphragms simultaneously compensate for the length of an element acting on the hydraulic compensation element and absorb a thermal change in volume of the fluid in the hydraulic compensation element.

"Hydraulically closed" is understood to mean that the hydraulic compensation element with an outside of the hydraulic compensation element located hydraulic system is not hydraulically connected.

The hydraulic compensation element can therefore advantageously be manufactured and tested as a modular unit.

The term "axial" is understood to mean the axis of the housing, which runs through its two end faces. "Axial" is also understood to mean the direction of the force effects and transmissions of the hydraulic compensation element. However, “axially” is also understood to mean “essentially axially”.

The hydraulic compensation element is preferably used in an injector and / or in a common rail injector.

The invention is explained in more detail using the following exemplary embodiments.

Show

FIG. 1 shows a first, simple design of the hydraulic compensation element,

Figure la the installation of the hydraulic compensation element in a metering device

FIG. 2 shows a development of the hydraulic compensation element in a side and cross-sectional view, in which the housing has an axial bore which leads to a hydraulic chamber via a check valve,

Figure 3 is a development of the hydraulic compensation element in a side and cross-sectional view in which the piston parallel to the housing and Piston axis has running bores which are connected to the hydraulic chamber via a check valve,

FIG. 4 shows a further development of the hydraulic compensation element in which the filling bore leads to a storage volume,

5 shows a development of the hydraulic compensation element in which the housing is extended by a sleeve and a second hydraulic chamber is formed between the housing and the piston,

6 shows a development of the hydraulic compensation element consisting of a combination of FIGS. 3 and 5,

FIG. 7 shows the functioning of the hydraulic compensation element according to FIG. 6,

8 shows a development of the hydraulic compensation element consisting of a combination of FIGS. 3 and 5, the filling bore opening into an axial bore of the housing,

9 shows a development of the hydraulic compensation element in which the housing expanded by a sleeve according to FIG. 5 has axial through-bores,

10 shows a development of the hydraulic compensation element in which both the housing expanded by a sleeve and the piston according to FIG. 5 have axial bores which are each connected to a check valve,

FIG. 11 shows a development of the hydraulic compensation element in which the housing is extended by a sleeve. tert, which has a smaller outer diameter compared to the sleeve according to FIGS. 5 to 10,

Figure 12 is a further development of the hydraulic compensation element with an additional membrane, and

FIG. 13 shows a further development of the hydraulic compensation element consisting of a combination of FIGS. 5 and 12

Figure 1 shows how the hydraulic compensation element 1 with the exception of a filling hole 11 is constructed rotationally symmetrical. The hydraulic compensation element 1 consists of a hollow cylindrical housing 2, a piston 3 slidably accommodated in the housing, elastic membranes 4, 5, which are each attached to the end faces S1 and S2 of the housing, storage volumes 8 and 10 which are between the membranes and the Housing are formed, a hydraulic chamber 9 and from the filling bore 11 and an associated closure element 12. The end regions of the piston 3 can protrude from the end faces of the housing. The hydraulic chamber 9 and the storage volumes 8 and 10 are filled with a fluid of selectable viscosity.

The piston 3 consists of a piston rod 3b and a region with an enlarged cross section 3a, which can also be referred to as a disk 3a. The disk 3a has two surfaces 3a 'and 3a' ', each of which has a pressure-effective effect on the end faces of the housing. The surfaces are preferably circular ring surfaces which are aligned essentially axially in order to achieve the axial pressure effectiveness. The orientations of the surfaces do not necessarily have to run parallel to the piston axis, but can be at an angle to the piston axis as long as their axial pressure effects are guaranteed. The

If necessary, surfaces can be made rough or at least uneven. The piston 3 preferably has one Bore 13 through which electrical leads K can be passed. The electrical feed lines can be used to control a piezo actuator in an injector.

The housing 2 has an inner space which has at least two different inner diameters. The inner wall of the housing also has at least one axially pressure-effective surface 2a ^Λ , which lies opposite the surface 3a of the piston 3. The surface 2a is also preferably an annular surface here.

A clearance is formed between the housing 2 and the piston 3, which is characterized by clearance fits 6, 7, the hydraulic chamber 9 and the storage volumes 8, 10. The outer contour of the piston 3 is adapted to the shape of the inner wall of the housing 2, but the above-mentioned latitude is always guaranteed in order to allow a relative displacement between the piston 3 and the housing 2. The play in the area of the storage volumes 8 and 10 also allows the bulges and contraction of the membranes 4 and 5.

Clearance fit and hydraulic chamber

The clearance fits 6, 7 seal the hydraulic chamber 9 formed from the upper, axially pressure-effective surface 3a of the piston 3 and the opposite axially pressure-active surface 2a of the housing 2 from the storage volumes 8, 10. A fluid exchange between the storage volumes 8, 10 and the hydraulic chamber 9 is only possible in a strongly throttled manner due to the clearance fits 6, 7. The storage volumes 8, 10 are preferably hermetically sealed to the outside by means of the elastic membranes 4, 5, which are sealingly attached to the piston 3 and the housing 2 by means of welding points 4a, 5a. The elastic membrane can therefore also be referred to as elastic seals. The membranes 4 and 5 are preferably annular flat membranes. The elastic membranes, which are designed identically in terms of their pressure-loaded area and their force transmission into the piston 3 on the small diameter and in the housing 2 on their outer diameter, form together with the piston 3 and the housing 2 a hermetically sealed interior of the hydraulic compensation element 1, which is degassed with silicone oil suitable viscosity is filled bubble-free under pressure. Such a structure is pressure balanced, ie no net force is created between the filling pressure

Piston and housing caused. The interior of the hydraulic compensation element 1 differs from the aforementioned interior of the housing 2 in that the interior of the hydraulic compensation element in particular also includes the storage volumes 8 and 10.

The hydraulic compensation element 1 is evacuated by means of a suitable device in order to then be filled with degassed silicone oil via the filling bore 11 without bubbles and under a basic pressure. A closure element 12 in the form of a ball or a dowel pin closes the filling hole 11 in a gas-tight manner. The closure element 12 can additionally be secured by gluing or welding. The filling hole 11 leads to the clearance fit 6, so that the parasitic fluid volume enclosed in the filling bore 11 (a basic principle of the hydraulic compensation element is to keep the enclosed volume as small as possible; therefore each extra volume is disruptive, i.e. parasitic), only throttled with the clearance fit 6 the hydraulic chamber 9 is connected. A hydraulically unthrottled or only slightly throttled connection of the filling bore 11 to the hydraulic chamber 9 instead resulted in a loss of rigidity in the hydraulic chamber.

The width and length of the clearances 6.7, and thus their throttling effects, are preferably precisely matched to the viscosity of the fluid. The use of silicone oil like Baysilone® M, for example, is preferred due to its significantly lower viscosity dependence on temperature compared to mineral oils in a relevant temperature range and the resulting simplified adjustment of the throttling effect of the clearance fits.

The function of the hydraulic compensation element 1 is based on the fact that, by means of the hydraulic chamber 9, very high forces can be transmitted briefly with a negligible relative displacement between the piston 3 and the housing 2, and the hydraulic chamber has a rigidity comparable to that of a solid body, while a slow one relative displacement between the piston 3 and the housing 2 takes place within certain limits, such as the height of the hydraulic chamber, almost without force (stiffness 0). The hydraulic compensation element can therefore be used in switching valves that operate for short or periodic periods, the time of force transmission of the hydraulic compensation element 1 preferably being very short compared to the emptying time of the hydraulic chamber 9. In principle, no static forces can be transmitted between the piston 3 and the housing 2 via the hydraulic chamber 9.

Figure la shows how the hydraulic compensation element is installed in a metering device D. The metering device consists of an actuator unit A, which has a piezo actuator P, a tubular spring C, and two end caps E1 and E2. The piezo actuator is preferably prestressed by means of the tubular spring connected at its two ends to the end caps E1, E2. The end caps E1, E2 are preferably each connected to the piston 3 of the hydraulic compensation element and to a valve needle VN of a valve unit B. The piston 3 of the hydraulic compensation element is preferably provided on its end faces S1 with a reset mechanism, for example with a spring FI, and on its second end face with the one end cap El of the actuator. connected ator unit. The housing 2 of the hydraulic compensation element is connected to a fixed bearing L and is therefore not movable relative to the actuator unit. More specifically: the housing 2 is with an inner wall Dj. _nn the housing of the metering device D, preferably by means of welds SN, connected. A force emanating from the piezo actuator is applied to the piston 3 of the end face S2, on which the storage volume 8 directly adjoins the disk 3a of the piston. The force from the piston 3 is transmitted to the housing 2 via the fluid of the hydraulic chamber 9, so that the hydraulic compensation element is rigid. Thus, the end cap E2 presses stiffly on the valve needle VN of the valve unit B so that a sealing element or a valve can be opened. With slow expansions of the piezo actuator, as described in more detail below, the piston can be displaced relative to the housing without force. The piston can be pushed back into a central position by means of the reset mechanism F. It is also possible for the end cap E1 to be firmly connected to the housing 2 on the end face S2 and for the piston 3 to be connected to the housing of the metering device. In this case, the spring Fl would be connected to the front side S1 of the housing.

It is preferred that a reset mechanism in the form of a stop which is designed as a valve seat or plate VS is provided in the valve unit. In this case, the valve needle VN could not be pulled up over the plane of the valve seat VS, so that the piston is forced to always return to a central position relative to the housing.

An alternative reset mechanism in the form of a stop can also be provided, which is implemented as a taper of the housing of the metering device. The end cap E2, which is preferably formed with a plunger, and the stop can each have structures which mutually represent a key-lock principle in such a way that during the manufacture of the metering device end cap E2 is guided past the stop and after its subsequent rotation the end cap can no longer be passed past the stop or pulled. A maximum distance between a sealing element of the valve and the end cap would thus always be guaranteed.

The preferably soft spring F2 is also a resetting mechanism, the net force effect of which causes the piston 3 to be pulled back into a central position after one or more forces have been applied.

If a pressure force F, which wants to move the piston 3 and the housing 2 axially towards one another, is transmitted to the housing 2 by the piston 3 via the fluid in the hydraulic chamber 9, the pressure in the hydraulic chamber 9 increases with an annular area A _k by ΔP = F / A _{k in} relation to the basic pressure P ₀ in the storage volumes 8, 10, and the height of the hydraulic chamber 9 begins to decrease slowly with a constant drift speed, the fluid in the hydraulic chamber 9 via the clearances 6.7 in the storage volumes 8 , 10 emptied into it. The volume flow of the fluid is directly proportional to ΔP, and thus to F, and inversely proportional to the throttling effect of the clearances 6.7. A power transmission can therefore only be understood as brief if the drift path of the piston or the volume flow of the fluid is negligible during the period of the force acting on the piston 3 and the hydraulic chamber 9 can be regarded as almost tight during the force acting.

The hydraulic compensation element 1 is also able to transmit tensile forces between the housing 2 and the piston 3, but the tensile force is limited by the basic pressure P ₀ prevailing in the hydraulic compensation element. A tensile force is understood to mean a force which wants to move the piston and the housing axially away from one another and thus to increase the volume in the hydraulic chamber. A pulling attack on the hydraulic compensation element Force, that is, for example, if the piston 3 is pulled on the end face S2, leads to a pressure drop in the hydraulic chamber 9 of ΔP ₀ = F / A. However, since pressures cannot become negative, the maximum conceivable pressure drop is ΔP = P _o and so that F _train _ _raax = P ₀ * A applies.

Fluid behavior with temperature fluctuations

The fluid volume enclosed in the hydraulic compensation element 1 is optimized to a minimum. This is necessary because in the automotive-relevant temperature range of -40 to 140 ° C, the operating temperature at a filling temperature of 20 ° C fluctuates between -60 ° C and 120 ° C. If the amount of fluid filled in at 20 ° C. was at a certain volume V ₀ , the volume of silicone oils, for example Baysilone, changes with a typical volume change coefficient of 1 * 10 ^"3 K ^_1 in a relevant temperature range by approx. -6% up to 12% based on the original volume V ₀ .

If the height of the hydraulic chamber 9 is 150 μm in the idle state of the piston 3 and the interior of the housing 2 has a large diameter of 14 mm and a smaller diameter of 5 mm, the enclosed fluid volume at 20 ° C. is approximately 36 mm ³ (V ₀ ) a fluctuation range from -2.2mm ³ at -40 ° C to 4.3mm ³ at + 140 ° C, again based on the original volume V _o at a room temperature of 20 ° C.

membrane

However, this thermal change in the fluid volume is advantageously compensated for by the elasticity of the membranes 4, 5. They bulge when the temperature rises and enlarge the interior of the hydraulic compensation element 1. A reverse effect occurs at a reduced temperature. The membrane is designed to be sufficiently pressure-stable and elastic so that the temperature-dependent volume Fluctuation of the fluid is compensated and at the same time the piston 3 and the housing 2 are axially displaceable relative to each other almost without force. The change in the fluid volume therefore does not lead to a net force effect between the piston 3 and the housing 2.

The diaphragms 4, 5 are optimized in such a way that they oppose an axial relative displacement between the piston 3 and the housing 2 only by a negligible force due to the lowest possible mechanical rigidity. They nevertheless form a stop which prevents the piston 3 and the housing 2 from being pulled apart from a position defined by the welding positions 4a, 5a of the membranes. If, for example, the membrane 4 rests on the housing 2 and / or the lower membrane 5 on the piston, further pulling apart is prevented.

The diaphragms are largely flat, with a natural curve in the direction of the housing or the piston. They preferably have a sealing function and can therefore also be referred to as elastic seals.

The elasticity of the membranes 4, 5 can be increased with the same compressive strength by layering d, n membranes of the thickness d / n instead of a single membrane 4, 5 with a wall thickness. By suitable shaping of the membrane area close to the piston 3, the mechanical stresses near the welding line 4a or 5a can be kept within material-compatible limits.

By shaping the membranes 4, 5 by means of an embossed structure, such as, for example, a wave structure in a radial orientation or a wave structure in the circumferential direction, the membrane can be optimized in terms of its rigidity against axial displacements of the piston and also in terms of its storage capacity. The membranes can also be optimized by means of a suitable profile in the membrane thickness, that is to say by varying the thickness, or seals can be designed, for example, in the form of bellows or metal bellows. The membrane can also be designed in multiple layers. Flexible seals can be designed as a combination of flat membranes with bellows. Plastic, rubber or fiber reinforced membranes are preferably used.

The material selection for the membrane is not limited to metallic materials. All types of sufficiently pressure-resistant and temperature-stable materials or composite materials can be used, e.g. elastomeric plastics or CFRP, GRP composite fiber materials.

The use of mechanically soft and low-friction sealing metal, plastic or rubber membranes avoids the use of friction-causing and therefore function-impairing but inexpensive O-rings of the prior art.

The diaphragms 4, 5 can be designed asymmetrically with regard to their pressure-loaded surface and their introduction of force into the piston 3 at the small diameter and into the housing 2 at their large diameter. In this case, the hydraulic compensation element 1 is no longer pressure-balanced, but instead develops a pressure-dependent force effect which can be advantageous for certain applications. In this case, an internal pressure in the hydraulic fluid that changes with the temperature leads to a net force effect between the piston 3 and the housing 2 that is directly proportional to the pressure. The proportionality factor can be set in a targeted manner.

In order to increase the bulging of the membranes, it is preferred to make them from a material with a different, preferably higher, temperature expansion coefficient than that of the housing and / or the piston. With increasing temperature the difference in the coefficients of thermal expansion is a greater increase in the diameter of the membranes 4, 5 compared to that of the housing 2 and / or the piston 3. The hermetically sealed fastening of the membranes to the housing 2 and the piston 3 creates a clamping condition on the inside and outside diameters of the membranes 4 ; 5. Therefore, the difference in the diameter increase between the housing 2 and the membrane can only be absorbed in an increased bulging of the membranes, which corresponds to a bimetal effect. If required, the membranes can also have different thermal expansion coefficients.

Different coefficients of thermal expansion of the housing and the piston

In the case of a hydraulic compensation element of the type presented, a low-friction movement of the piston 3 relative to the housing 2 must be ensured, since otherwise its intended compensating function would not exist or would only be possible to a limited extent. For this purpose, the fit dimensions and tolerances of the piston and housing must be selected so that there is positive play. For a low-friction and low-jerk movement between the piston and the housing, there is also a sufficient surface area on the outside of the piston and / or the inside wall of the housing, in particular a low surface roughness, as can be produced, for example, by grinding, and to avoid tilting. a sufficient lead length, advantageous. Adherence to the fit dimensions of the piston and cylinder is ensured in such a way that not only in the assembled state but also in the stationary and unsteady operation of the hydraulic compensation element, there is no jamming or frictional sliding (stick-slip) of the piston in the housing, for example due to a stronger one thermal expansion of the piston in relation to the housing or a stronger thermal contraction of the housing in relation to the piston. Especially in non-stationary operation and at high operating frequencies arise due to the high and rapidly changing heat release of the actuator with simultaneous cooling by the fuel, radial temperature gradients, which can lead to different thermal expansion of the piston and cylinder and, if not properly designed, clamping. This can be prevented by the following measures:

a.) the piston and the housing are made of the same material or materials with the same thermal expansion coefficient. To avoid pinching, a sufficiently large gap between the piston and cylinder in the range of 10 to 50 μm in connection with a fluid with a higher basic viscosity in the range of 100 to 1000 centistokes should be selected with a sufficient guide length of the piston in the housing to prevent tilting.

b.) Does the piston heat up e.g. stronger than the housing due to a connected drive element such as due to the piezo actuator (this creates a radial temperature gradient that should not be neglected), a material with a smaller thermal expansion is selected for the piston 3, so that the piston does not start to jam in the narrow clearance fits 6.7.

c.) If one can assume that the piston 3, the hydraulic fluid and the housing G are always at almost the same temperature, the temperature influence on the gap flow between the clearance fits 6, 7, 19 can be in the state loaded by an actuator of the hydraulic system can be compensated in a wide range if the piston has a suitably chosen higher thermal expansion than the housing. The explanation is that the viscosity of the hydraulic fluid decreases with temperature according to an exponential law and the volume flow of the hydraulic fluid increases exponentially along the clearance fits. The volume flow is proportional to the 3rd power of the width of the Game fits, which can also be called fit dimensions. The fit diminishes linearly with the temperature and thus the temperature effects on the fit and on the viscosity are opposite.

Compensation for a thermally induced change in length of the piezo actuator

Effects with sufficiently different time scales occur with injection valves or injectors for internal combustion engines:

Typical injection times of approximately 1 ms to 3 ms in which the hydraulic chamber 9 has to transmit a force.

Slow thermal processes in the range of seconds to minutes, whereby fluid exchange in the hydraulic compensation element must take place without hindrance.

However, the invention provides that a piezo actuator A with the

Piston 3 or connected to the housing 2 and its slow thermally induced change in length does not lead to a power transmission between the piston and the housing. When the piezo actuator A warms up and this leads to its change in length in the range of seconds or minutes, the change in length only causes an almost negligible force on the piston 3 or on the housing 2. The injector function is therefore activated by slow e.g. thermal changes in length are not affected. However, impulsive force loads on the hydraulic compensation element can still be transmitted stiffly.

Repeatedly acting forces on the hydraulic compensation element 1 accumulate, however, by adding the displacement path between the housing 2 and the piston 3. For this reason, a reset mechanism as described for FIG Blanking gap returned to its starting position. The blanking interval is the time in which the hydraulic compensation element is not loaded with a force and the piston can move back into its rest position.

Since a fluid valve to be driven is almost always realized in the form of a plate or seat valve in an injector, it has at least one fixed stop in the form of a sealing surface or sealing line VS, by means of which the closed initial position of the closing element is fixed relative to the valve housing. If the hydraulic compensation element 1 is used in a valve of this type, in which on the one hand the piston 3 on the end face S1 is firmly connected to the piezo actuator A and on the end face S2 with the closing element like a valve needle, and on the other hand the housing 2 of the hydraulic compensation element 1 is connected to a fixed bearing B like a valve housing, a reset mechanism is provided in the simplest way.

The resetting can also be realized, as shown in FIG. 1 a, by a spring FI acting on the piston 3, which holds the piston 3 in contact with the piezo actuator and the piezo actuator in contact with the valve needle. Here, the hydraulic compensation and drive elements as well as the drive element and closing element do not have to be firmly connected but only kept in contact.

At higher duty cycles it can happen that the pressure in the hydraulic chamber at the end of the blanking interval has not yet reached the value PO of the rest position of the piston (incomplete refilling) and a dynamic tensile force is established between the piston 3 and the housing 2. An analogous behavior occurs with repeated tensile forces. The duty cycle is defined by the ratio of the time in which a force is transmitted via the hydraulic chamber 9 to the duration between two successive forces (period duration). However, the full backflow of the fluid can also be achieved at higher pulse rates by means of a check valve 15 as presented below.

Presentation of the Flappervalves and the holes required for the return

FIG. 2 shows a check valve 15 accommodated in the hydraulic chamber 9, hereinafter also referred to as "flapper valve", which covers at one end an additional bore 14 incorporated in the housing 2 and connecting the storage volume 10 with the hydraulic chamber 9. The flapper valve 15 effects an adjustment of the fluid flow through the bore 14. It is fixed by means of weld points 17 on the annular surface 2a ^{λ of} the housing 2 so that it can be swung open relative to the axially pressure-effective annular surface of the piston 3a.

If the piston 3 is loaded by a force F (in the direction of the arrow), an overpressure arises in the hydraulic chamber 9, so that the flapper valve 15 remains pressed against the housing 2 and the bore 14 is closed. This keeps the hydraulic compensation element stiff. As soon as the force F decreases in the blanking gap, the membranes 4, 5 or generally a resetting mechanism cooperating with the piston 3 push the piston 3 again in the direction opposite to the force F. This creates a negative pressure in the hydraulic chamber 9, which forces the flapper valve to open. The backflow of the liquid from the storage volume 10 into the hydraulic chamber 9 is accelerated, since the backflow is possible not only through the clearance fits 6, 7, but also through the additional bore 14.

The lower drawing of this figure is a cross-sectional view of the hydraulic compensation element, in which an annular flapper valve 15 is shown. FIG. 3 shows how, instead of the housing 2, the piston 3 is provided with bores 16 which also connect a storage volume 8 to a hydraulic chamber 9. If a vacuum occurs in the hydraulic chamber 9 at the beginning of the blanking interval mentioned above, and if the flapper valve 15, which in this exemplary embodiment is attached to the annular surface 3a ^{λ of} the piston 3 in a hinged manner, is drawn through the differential pressure, the bore 16 released and the fluid flow through the bore into the hydraulic chamber is significantly accelerated.

The lower drawing of Figure 3 is also a cross-sectional view of the upper drawing.

The hydraulic compensation elements according to FIGS. 2 and 3 with their respective flapper valves achieve duty cycle ratios of up to 0.8.

FIG. 4 shows a development of a hydraulic compensation element 1 according to FIG. 1, in which a filling bore

11a is connected to the storage volume 10. In contrast to FIGS. 1 to 3, the clearance fit 6 throttles the entire axial length in this embodiment, so that its throttling effect is increased.

Hydraulic compensation element with two hydraulic chambers

FIG. 5 shows a hydraulic compensation element in which the housing 2 is expanded by a part 18. The enlarged part 18 can be referred to as a sleeve. The outer contour of the piston 3 is again adapted to the inner wall of the sleeve 18 such that an additional clearance 19 is formed between the piston 3 and the sleeve 18. The sleeve 18, the piston 3 and the housing 2 define a second un- tere hydraulic chamber 20, which is * of the sleeve 18 and an axially pressure-effective surface 3a ^Λ formed between an axially pressure-effective surface 2a of the piston. 3 The GE- housing 2 and the sleeve 18 are fluid-tight and mechanically rigid, for example by means of welds 21 connected to each other. The lower membrane 5 is hermetically sealed at its inner diameter at the lower end of the piston 3 and at its outer diameter on the sleeve 18.

A structure with two hydraulic chambers 9 and 20 leads to a doubling of the rigidity of the hydraulic compensation element when it is loaded by a brief force. The doubled stiffness is advantageously achieved with almost the same outer diameter as the outer diameter of the hydraulic compensation elements with only one hydraulic chamber 9.

By means of the two hydraulic chambers 9, 20, a force can be applied to one of the two end faces S1 or S2. The hydraulic compensation element is thus bidirectional. A force can be applied to the housing 2 or, alternatively, to the piston 3 without the compensation and force transmission properties of the hydraulic compensation element 1 being changed. The housing 2 expanded by the sleeve 18 can on the one hand be rigidly connected to a fixed bearing B and on the other hand the piston 3 can be connected to a piezo actuator A. Conversely, the housing 2 could be connected to the piezo actuator A and the piston 3 could be rigidly connected to the fixed bearing B. his. Only the relative movement between the piston 3 and the housing 2 is decisive.

The functioning of the hydraulic compensation element 1 according to FIG. 5 can be described in more detail as follows:

If a force F acts on the piston 3 on the end face S2 through the piezo actuator A, and if the hydraulic compensation element 1 is attached to the housing 2, for example on a fixed bearing B, then a pressure Pl is established in the hydraulic chamber 9 with chamber height h1 which is only affected , At the same time, in the second hydraulic chamber 20 with chamber height h2 a pressure P2 on. Both hydraulic chambers 9, 20 have hydraulically active circular ring surfaces 2a, 2a ^Λ> , 3a ', 3a with the respective surfaces A ^ stretched between the inner clearance fits 6, 19 and the outer clearance fit 7. Hence

F = -A _k {P ₂ -P _x )

If the fluid has a compressibility K, the following applies

and ^ o- K h,

where Po is the current load-free basic pressure on the fluid as it prevails in the storage volumes 8, 10 and x denotes the displacement of the piston 3 caused by the force F. So it applies

so that the spring rate C of the hydraulic chambers 9 and 20 as

can be expressed.

Due to the pressure change in the hydraulic chambers 9, 20 with a force F, the height of the hydraulic chamber 9 begins to decrease at a constant speed and, on the other hand, the height of the hydraulic chamber 20 begins to decrease with a constant to increase speed, the fluid being emptied from the hydraulic chamber 9 into the hydraulic chamber 20 via the clearance fit 7. In addition, owing to the pressure drop in the hydraulic chamber 20 compared to the pressure P ₀ prevailing in the storage space 8, fluid flows from the storage space 8 into the hydraulic chamber 20.

It can be assumed that there is always an increase in pressure in the first hydraulic chamber and a decrease in pressure in the second hydraulic chamber at the same time. Since a pressure reduction is only possible to a minimum up to the vapor pressure of the fluid (close to 0 bar), the hydraulic chamber in which the pressure reduction takes place becomes soft with further load, so that from this moment the relation

C = κ holds

If, on the other hand, the hydraulic compensation element on the housing 2 is loaded with a force from the other end face S1, the following relations apply:

P = P - ^■ κ and

P ₂ = P ₀ + K- h,

The hydraulic chambers 9 and 20 thus swap their roles in the sense that there was a pressure drop in the hydraulic chamber 9 in the previous pressure rise and vice versa. To the outside, however, the hydraulic compensation element behaves identically regardless of the direction of force; so it's bidirectional. When the piston 3 or the disk 3a of the piston 3a is in an intermediate position, where the hydraulic chamber heights h1 and h2 are the same, the rigidity of the hydraulic compensation element 1 is doubled compared to a hydraulic compensation element with only one hydraulic chamber. According to the formulas given above, the stiffness is even higher in the off-center position.

FIG. 6 shows the combination of a piston 3 provided with axial bores 16 with a housing 2 expanded by a sleeve 18. As already explained, the bores 16 can interact with a flapper valve 15. The advantages of a piston 3 additionally provided with bores 16 are, as mentioned in the description of FIGS. 2 and 3, combined with the advantage of increased rigidity with two hydraulic chambers.

FIG. 7 shows the mode of operation of the hydraulic compensation element presented in FIG. 6. If, as shown on the left-hand side, the hydraulic compensation element on the piston 3 is subjected to a force F, an overpressure arises in the hydraulic chamber 9 in relation to the lower hydraulic chamber 20 and this pressure difference keeps the flapper valve 15 tightly closed. If, as on the right-hand side of the figure, the pressure conditions are reversed and a tensile force acts on the piston 3, the flapper valve 15 opens the bores 16 in order to enable an almost unthrottled fluid flow f from the lower hydraulic chamber 20 into the upper hydraulic chamber 9, whereby the Piston 3 is easily displaceable. The tractive force can be achieved, for example, by means of the reset mechanism already mentioned, which pulls the piston 3 back or pushes it into a defined starting length after an application of force and the drift caused thereby.

FIG. 8 shows a hydraulic compensation element 1 according to FIG. 5 which has bores 14 and 14a in the housing 2 and the filling bore 11 with at least one of these bores, for example, connected to the bore 14a. The bores 14 and 14a still connect a hydraulic chamber with a storage volume. The flapper valve 15 can close the bores 14 and 14a as shown in FIG. The advantages of the holes 14 and 14a for a quick backflow of the

Fluids can thus be combined with the filling bore 11a connected to the storage volume 10 (the clearance fit 6 throttles over its entire length) according to FIG. 4.

FIG. 9 shows how the sleeve 18 from FIG. 5 is provided with bores 14b. When a force F is applied, a negative pressure is created in the hydraulic chamber 20 so that the hydraulic suction resulting therefrom presses the flapper valve 15 against the axially pressure-effective circular surface 18a of the sleeve 18 and the bores 14b are closed. The fluid can therefore only flow out of the storage volume 8 through the clearance fit 19. In the blanking interval, the piston 3 moves again in the direction of the piezo actuator A, so that the flapper valve 15 is opened and the return flow of the fluid back into the storage volume 8 is facilitated.

The filling hole 11 is machined into the housing 2, but it can also be machined into the sleeve 18. If required, it can also open into the bore 14b and thus also be connected to the hydraulic chamber 20 and the storage volume 8.

FIG. 10 shows a further development of the invention, particularly suitable for the case where the hydraulic compensation element 1 is rigid under the application of force, but should prove to be soft under tensile forces, and both hydraulic chambers 9, 20, on the one hand, and on the other hand with the storage volumes 8 , 10 let the fluid flow back slightly. In this variant, the hydraulic compensation element is no longer bidirectional due to the flapper valves, only the rigidity is increased in the direction of force indicated by the arrow. Both the housing 2 and the piston 3 as well as the sleeve 18 are thus provided with holes 14 and 16 or 14, 14a, 14b, and 16. Each hole is provided with a flapper valve 15 at least on one side. Two Flappervalves 15 are received in the hydraulic likkammer 9 respectively to the axial pressure-effective surfaces 2a and 3a ^Λ of the housing 2 and the piston 3 are fastened, and 14 and 16 cover the holes 14a. If the piston 3 is loaded by a force F, the flapper valve 15 attached to the annular surface 18a of the sleeve 18 remains closed due to the (suction) pressure drop that has occurred in the hydraulic chamber 20. In the hydraulic chamber 9 there is an overpressure which closes the flapper valve 15 on the housing 2 and the opposite flapper valve 15 on the piston 3. The hydraulic compensation element is thus rigid and fluid movement is only possible through the clearance fits 6, 7, and 19. In the blanking gap, all the flapper valves 15 open because the vacuum in the blanking chamber in the hydraulic chamber 9 pulls both flapper valves in its volume inside and the pressure that rises again in the hydraulic chamber 20 both the flapper valve 15 attached to the piston 3 and the flapper valve 15 attached to the sleeve 18 also opens.

FIG. 11 shows how the hydraulic compensation element presented in FIG. 5 is realized with the smallest possible cross section. The housing 2 is expanded by a housing part 18b, which is welded firmly to the original housing 2 without, in contrast to the sleeve 18 according to FIGS. 5 to 10, this expansion leading to an overall enlarged cross section of the expanded housing. As with the sleeve 18, the sleeve 18b can have bores for the same purpose and can be provided with flapper valves 15. In order to achieve a minimal cross-section of the entire device, it is preferred to countersink the welding points 21a or at least to carry them out smoothly on the surface of the housing. In manufacturing, this can be achieved by smoothing the welds. FIG. 12 shows a development of the invention in which the hydraulic compensation element 1 is designed such that a compensation store AS is provided on the outside of the housing 2 and spans it. The compensating accumulator has an elastic membrane 22 and a storage volume 23 enclosed underneath. The storage volume 23 is formed between the outside of the housing 2 and the elastic membrane 23 and is fluidly unrestrictedly connected to a further storage volume 10 via an additional bore 11b. The storage volume 8 can preferably also be connected to AS unthrottled. The elastic membrane is preferably an elastic sleeve which spans the outside of the housing 2, that is to say its outer surface.

Due to a rise in temperature, the oil filling in the hydraulic compensation element expands. If the increase in volume is not compensated for (compensated), the oil pressure rises in the worst case until plastic deformation and destruction of the membranes 4,5. The elastic sleeve provides a temperature-dependent collection volume, so that a dangerous pressure rise is avoided or the oil pressure is ideally kept constant. The elastic sleeve 22 is welded at its edges at the welding points 22a to the outside of the housing 2 and preferably has metal and / or plastic parts. The size or concentration of the respective proportions is matched to the desired elasticity of the elastic sleeve.

The compensation store AS serves in particular to compensate for the thermal expansion of the fluid located in the hydraulic compensation element and is particularly advantageous with regard to the largest possible working temperature range of the hydraulic compensation element, such as, for example, when it is installed in an injector. 200303122

28 In the manufacture of a hydraulic compensation element 1 provided with the compensation accumulator AS, the elastic sleeve 23 with a suitable wall thickness, length and suitable coefficient of thermal expansion is pushed tightly over the housing 2 and welded to it.

In order to be able to provide sufficient space for the expansion of the elastic sleeve 22 of the compensating _accumulator AS between the housing 2 of the hydraulic compensation _element and the inner wall D _{ιnn of} the housing of the metering device D, it is preferred that the housing 2 of the hydraulic compensation element by means of a spacer with the inner wall D _ιrm the metering device is mechanically connected.

FIG. 13 shows how a bidirectional hydraulic compensation element 1 according to FIG. 5 is also provided with the elastic sleeve 22. The elastic sleeve is welded to the housing 2 and with its enlarged part 18b via the weld seam 22a. However, the compensation accumulator AS is preferably not placed over the filling hole, since otherwise the hydraulic compensation element would not be easily filled. The weld seam 21 is preferably flush with the outer contour or recessed so that the stretching of the elastic sleeve 22 is not hindered.

The compensation memory AS with its elastic sleeve 22 can of course be used for all of the exemplary embodiments of this invention mentioned here.

In a development of the invention, the hydraulic compensation element 1 can be connected to an external hydraulic accumulator (not shown). The interior of the hydraulic compensation element would be connected to the external hydraulic accumulator via a bore through the housing 2, wherein this hydraulic accumulator can be part of a metering device or generally a device of a motor vehicle engine. The additional hydraulic accumulator was a further increase in the lumen compensation, which is realized according to Figures 12 and 13 by the compensation memory AS with the elastic sleeve.

Claims

claims

1. Hydraulic compensation element, which has a closed hydraulic system and can be filled with a fluid, comprising:

- a hollow cylindrical housing (2) with an interior which has at least two different transverse axial diameters and - the inner wall of the housing has at least one axially pressure-effective surface (2a ^λ ),

- A closable filling hole (11) leading into the interior of the housing, a piston (3) accommodated axially displaceably in the housing with - at least one axially pressure-effective surface (3a ^Λ ) which is opposite the axially pressure-effective surface of the housing and has an outer contour that corresponds to the The inner wall of the housing is adapted in such a way that some clearance fits (6, 7) with a fluid-restricting effect are formed between the piston and the housing

- At least one hydraulic chamber (9), which is formed between the at least one axially pressure-effective surface of the piston and the at least one axially pressure-effective surface of the housing and is connected on both sides to the clearance fits, elastic membranes (4, 5, 22) which are attached to the housing in such a way that storage volumes (8, 10, 23) are formed between the elastic membranes and the housing, with at least one elastic membrane (4, 5) connecting the housing to the front on each end face (S1, S2) of the hydraulic compensation element Piston connects.

2. Hydraulic compensation element according to claim 1, with two hydraulic chambers (9, 20), each between a first axially aligned surface (3a) of the piston and a first axially aligned surface (2a ^λ ) of the housing and are formed between a second axially pressure-effective surface (3a ' ^λ ) of the piston and a second axially pressure-effective surface of the housing (2a ^Λ, >. \

3. Hydraulic compensation element according to one of claims 1 or 2, wherein the housing (2) has at least one bore (14) which connects the at least one hydraulic chamber (9, 20) and the at least one storage volume (8, 10, 23) and which is connected at one end to a check valve (15) with which the fluid flow through the bore can be adjusted.

4. Hydraulic compensation element according to one of the preceding claims, wherein the piston (3) has at least one bore (16) which connects the at least one hydraulic chamber (9, 20) and the at least one storage volume (8, 10, 23) and which one One end is connected to a check valve (15) with which the fluid flow through the bore (16) can be controlled.

5. Hydraulic compensation element according to one of claims 3 or 4, wherein the check valve (15) is hinged to an axially pressure-effective surface (2a 3a ^λ ) of the piston (3) and / or the housing (2) is attached.

6. Hydraulic compensation element according to claim 4, wherein the check valve (15) is an annular plate.

7. Hydraulic compensation element according to one of the preceding claims, in which the filling bore (11a) is connected to a storage volume (8, 10, 23).

8. Hydraulic compensation element according to one of the preceding claims, wherein the elastic membrane (4, 5, 22) are elastic metal or plastic membranes.

9. Hydraulic compensation element according to one of the preceding claims, in which the elastic membrane (4, 5, 22) are fiber-reinforced.

10. Hydraulic compensation element according to one of the preceding claims, in which the elastic membrane (4, 5, 22) are designed in multiple layers.

11. Hydraulic compensation element according to one of the preceding claims, wherein the elastic membrane (4, 5, 22) have a thickness variation.

12. Hydraulic compensation element according to claim 9, in which the elastic membrane (4, 5, 22) are each biased differently.

13. Hydraulic compensation element according to one of the preceding claims, wherein the elastic membrane (4, 5, 22) have thermal expansion coefficients that differ from those of the piston and / or the housing.

14. Hydraulic compensation element according to claim 13, wherein the elastic membrane (4, 5, 22) each have different thermal expansion coefficients.

15. Hydraulic compensation element according to one of the preceding claims, wherein the piston (3) and the housing (2) have different thermal expansion coefficients.

16. Hydraulic compensation element according to one of the preceding claims, in which by means of the elastic membrane (4, 5, 22) a thermal volume change of the fluid located in the hydraulic compensation element can be absorbed.

17. Hydraulic compensation element according to one of the preceding claims, in which an elastic membrane (22) is designed as a sleeve and is attached to the outer surface of the housing (2).

18. Hydraulic compensation element according to one of the preceding claims, wherein the piston (3) has an axial bore (13) for the passage of electrical leads (K).

19. Hydraulic compensation element according to one of the preceding claims, in which through its clearance fits (6, 7) and hydraulic chambers (9, 20) a high rigidity compared to short-term force loads can be achieved and through the elastic membranes (4,5,22) a length compensation element acting on the hydraulic compensation element can be reached and a thermal volume change of the fluid located in the hydraulic compensation element can be absorbed.

20. Use of a hydraulic compensation element according to one of claims 1 to 19 in an injector.

21. Use according to claim 20 in a common rail diesel injector.

22. Dosing device which comprises a hydraulic compensation element (1) according to one of Claims 1 to 19 and an actuator unit (A), the actuator unit comprising at least one piezo actuator (P) and the piezo actuator being connected to the hydraulic compensation element.

23. Dosing device according to claim 22, in which the piezo actuator (P) is provided at both ends with end caps (El, E2) and one end cap (E2) with a valve needle (VN) and the other end cap (El) with the Piston (3) of the hydraulic compensation element is connected.

24. Dosing device according to one of claims 22 or 23, in which the piezo actuator (P) is prestressed by means of a tubular spring (C) which is firmly connected at both ends to the end caps (E1, E2).

25. Dosing device according to one of claims 23 or 24, in which electrical connections (K) are guided through an axial bore (13) of the hydraulic compensation element and through bores of the end cap (El) to the piezo actuator (P).

26. Dosing device according to one of claims 22 to 25, in which the housing (2) of the hydraulic compensation element is fixed, in particular by means of weld seams (SN), with a fixed bearing (L) and is not movable relative to the piezo actuator (P) is.

27. Dosing device according to one of claims 22 to 26, in which the piston of the hydraulic compensation element is connected at its one end cap (El) to a reset mechanism.

28. Dosing device according to claim 27, wherein the reset mechanism is a spring (F).

29. Dosing device according to one of claims 22 to 28, wherein the resetting mechanism is a stop in the form of a valve seat VS.

30. Dosing device according to one of claims 22 to 29, wherein the reset mechanism is implemented as a stop in the form of a taper of the housing of the dosing device.

1. Dosing device according to claim 30, in which the end cap E2 and the stop each have structures which mutually represent a key and lock principle.