EP2156047B1 - Fuel injector having low wear - Google Patents

Description

State of the art

The invention is based on known fuel injectors for injecting fuel into a combustion chamber of an internal combustion engine. These fuel injectors may in particular be fuel injectors for high-pressure accumulator injection systems, which are also referred to as common rail injectors. In this case, fuel from a high-pressure accumulator (rail) is supplied to the fuel injectors via a high-pressure inlet, and the injection of the fuel, i. the opening of the fuel injector, is usually controlled by an actuator, such as a magnetic or piezoelectric actuator.

One advantage of the common-rail injectors is that the injection behavior can be controlled very precisely via the actuator, so that even complex, for combustion particularly favorable injection curves can be realized. As a result, internal combustion engines with extremely low pollutant emission can be realized.

However, a problem occurring in common-rail fuel injectors in practice unpleasant phenomenon is the problem of pressure oscillations. Pressure oscillations in the common rail system can have undesirable or damaging effects on the operating behavior of the overall system and of the internal combustion engine. In addition, these can also reduce the durability of the overall system and individual injectors. Thus, pressure oscillations in the high-pressure region with an amplitude of about 300 bar are in many cases a commonplace event.

However, in practice, the pressure oscillations lead to nozzle seat wear and to an inaccurate quantity, in particular in the case of multiple injections. Therefore, it is attempted to counteract these pressure oscillations with damping measures, such as throttles and flow control valves.

An example of such a measure for the at least partial elimination of pressure oscillations is in DE 103 07 871 A1 disclosed. In the process, a high-pressure line will break proposed an injector and a common rail, which degrades the pressure waves generated during actuation of the injector in whole or in part. The high-pressure line has a first section and a second section, which are connected in parallel with one another.

In the DE 103 07 871 A1 shown arrangement contributes to the improvement of the vibration behavior and to the elimination of pressure oscillations. Nevertheless, the leaves DE 103 07 871 A1 Room for further improvements, especially since the proposed high-pressure line is structurally relatively complex and attenuates only pressure oscillations within a narrow frequency range due to the fixed length ratios usually.

Another common rail injector is out EP 1 612 401 A known.

Disclosure of the invention

The invention is based essentially on the idea of providing at least one further, separate vibration damping bore in addition to a high-pressure bore which is usually present in common-rail injectors and extends substantially parallel to the injector axis. This vibration damping bore comprises a damping channel extending essentially parallel to the injector axis in the injector housing. By "substantially parallel" are here and below also slight deviations from the parallelism to understand, preferably deviations of not more than 20 °, more preferably not more than 5 °.

In contrast to DE 103 07 871 A1 It is thus proposed to displace the vibration damping element into the injector body itself. The additional hole can be realized without major design effort and is therefore also easy to implement in terms of manufacturing technology.

Furthermore, the vibration damping behavior according to the invention is not based, as in DE 103 07 871 A1 , on a superposition of back and forth waves. As a result, an attenuation within a wide frequency range can be realized. The damping behavior is based on known systems rather on a "parallel circuit" of a damping channel, with a corresponding resistance to pressure waves and a corresponding volume of liquid.

It is provided that the vibration damping bore is located between a high-pressure region, in particular a high-pressure chamber, which is adjacent to the control chamber (for example, a high-pressure annular space which surrounds the control chamber), and the nozzle space. For example, in this case, the vibration damping bore may be substantially equal in length and extend substantially parallel to a high pressure bore which carries fuel under high pressure into the nozzle space to be injected from there through injection openings in the combustion chamber.

The throttle element comprises a first throttle element connecting the damping channel and the high-pressure annulus. Furthermore, the throttle element comprises a second throttle element connecting the damping channel and the nozzle chamber. This implementation of the throttle elements is also comparatively easy to implement in terms of production since in this case, for example, only one bore extending substantially parallel to the injector axis must be implemented for the damping channel and two bores for the throttle elements running obliquely thereto.

It has proven to be particularly advantageous for the damping properties, when the throttle elements have a cross section with a diameter between 0.1 and 0.8 mm, in particular between 0.2 and 0.5 and particularly preferably at 0.3 mm. For the damping channel, cross-sections of at least 1.0 mm, in particular at least 1.5 mm and particularly preferably of approximately 2.0 mm have proved to be advantageous. The latter value represents a compromise which can be realized in practice between the highest possible volume of liquid for the damping (inertial mass) and the space requirement of the damping channel in the injector body.

To be able to provide the highest possible damping volume as an "inertial mass", it is preferred if the damping channel has a length of at least 40 mm, in particular of at least 60 mm and particularly preferably of 90 mm. Thus, the damping channel can be implemented in today commercial fuel injectors, without their outer dimensions would have to be changed.

The fuel injector in one of the embodiments described above allows vibration damping of pressure oscillations within a wide frequency range. Thus, the wear of the fuel injector can be significantly reduced, and the life of the common rail systems can be significantly improved.

Brief description of the drawings

Embodiments of the invention are illustrated in the drawings and explained in more detail in the following description.

Figur 1

einen Druck- und Kraftverlauf für ein Common-Rail-System;

Figur 2

eine Systemantwort im Frequenzraum eines Common-Rail-Systems auf eine eigene Einspritzung und eine Nachbareinspritzung;

Figur 3

einen kumulierten Verschleiß für ein Ein-Injektor-System und ein gesamtes Common-Rail-System mit mehreren aktiven Injektoren;

Figur 4

ein Ausführungsbeispiel eines Common-Rail-Injektors gemäß der vorliegenden Erfindung;

Figur 5

ein Übertragungsverhalten eines Common-Rail-Injektors gemäß dem Stand der Technik im Vergleich zu einem erfindungsgemäßen Common-Rail-Injektor und

Figur 6

simulierter Verschleiß von Common-Rail-Injektoren mit verschiedenen Verschleiß reduzierenden Maßnahmen.

Show it

FIG. 1

a pressure and force curve for a common rail system;

FIG. 2

a system response in the frequency space of a common rail system to its own injection and a neighboring injection;

FIG. 3

cumulative wear for a one-injector system and an entire common rail system with multiple active injectors;

FIG. 4

an embodiment of a common rail injector according to the present invention;

FIG. 5

a transmission behavior of a common rail injector according to the prior art compared to a common rail injector according to the invention and

FIG. 6

simulated wear of common rail injectors with different wear reducing measures.

Based on FIGS. 1 to 3 First, wear effects in fuel injectors and their causes are to be outlined. The knowledge gained thereby provide the basis for a fuel injector according to the invention, which subsequently in the FIGS. 4 to 6 will be explained with reference to an embodiment.

In the FIG. 1 is a characteristic performance of a commercially available common rail injector shown as a time course. The upper curve (reference numeral 110) shows the pressure curve in bar, and the lower curve (reference symbol 112) the force, the nozzle needle or the injection valve member (hereinafter referred to as needle force), plotted in Newtons. Shown is the time course over a full injection cycle of a six-cylinder engine.

It can be seen that the operating behavior of the injection system can be divided into two areas: on the one hand into the area (in Fig. 1 denoted by reference numeral 114), in which the response to the own injection dominates, and in a second region (in Fig. 1 denoted by reference numeral 116), in which the influence of the neighboring islands predominates. In Fig. 1 The boundary between the two areas is approximately 0.027 s.

Within the first region 114, the vibration behavior, for example the vibration of the needle force 112, is dominated by the injection of the fuel injector under consideration. In contrast, in the second region 116, the vibration behavior is triggered by neighboring injections, that is to say by injections of adjacent fuel injectors of the internal combustion engine. This area of unfamiliarized vibrations is commonly referred to as the "telephony" area or "telephony" area.

The telephony effects, ie effects of the vibration in a fuel injector, which are triggered by neighboring injections, have a massive influence on the wear of the fuel injectors. Investigations have shown that the own injection only causes a total load per cycle of 1/3 of wear, in contrast to the influence of the neighboring injections, which have a share of about 2/3 of the wear.

In Fig. 2 is a possible cause of the effects described above shown. The gradients of the system responses, ie the derivatives I of the needle force (see curve 112 in FIG Fig. 1 ) in frequency space in arbitrary units. While curve 118 represents the system response to its own injection, curve 120 shows the system response to neighbor injections.

It can be clearly seen that essentially two main frequencies are excited both in the case of the own injection 118 and in the case of neighboring injections 120, namely a first frequency at approximately 650 Hz and a second frequency at approximately 3.2 kHz. These natural frequencies can be seen in both curves 118, 120 as steep peaks (maxima).

In addition, the system response 120 of the neighboring injections, ie the telephony effects, shows a multiplicity of further excitation frequencies. This means that the nozzle needle or the injection valve member of a fuel injector reacts very sensitively to excitations by neighboring injections and has a significantly more pronounced oscillation behavior.

This increased sensitivity, ie more pronounced system response or force gradient compared to a rail-side excitation by neighboring injections has a significant influence on the wear behavior of the fuel injectors. True, as from curve 112 in Fig. 1 shows that the vibration amplitudes, which are caused by neighboring injections, smaller than the vibration amplitudes in the range of their own injections 114. This might initially suggest lower wear. Decisive for the wear, however, are the gradients of these forces, which in Fig. 2 are shown and characterize the wear performance. Due to these higher gradients due to the telephony effects, the neighbor injections 120 have a higher proportion in the accumulated calculated wear values.

For the calculation of the wear different physical models can be generated, which will not be discussed in detail here. In Fig. 3 the result of such a wear model is illustrated graphically. In this case, the Y-axis 122 shows the accumulated wear, whereas the X-axis 124 symbolizes the time. The time is plotted over a complete injection cycle of a six-cylinder engine. In each case symbolically the injections of the individual injectors are designated by the reference numeral 126.

Again, two curves are shown: a curve 128, in which only the own injection is taken into account, and a wear curve 130 in which also adjacent injections are taken into account, which thus symbolize a fully activated system of an internal combustion engine.

Curve 128, in which only a single injector is active during an injection cycle, shows the expected, substantially flat course. Curve 130, in which the neighboring injections are also taken into account, however, shows a characteristic step pattern in which wear occurs not only in the own injection at the beginning of the cycle, but also in every subsequent injection by adjacent injectors. This is the result of in Fig. 2 shown system response of the non-activated injector to injections by Nachbarinjektoren.

As described above, one significant approach of the present invention is to reduce the sensitivities of a fuel injector to higher frequencies through geometric changes in the high pressure fluid system. Accordingly, a measure is proposed which is able to minimize the transmission behavior of the input pressure on the needle force in terms of its gain with respect to the natural frequencies.

In Fig. 4 an embodiment of a fuel injector 132 according to the invention is shown. The fuel injector 132 has an injector housing 134, which is modularly composed of a plurality of modules held together by a union nut 136.

In a Dusenmodul 138, an injection valve member 140 is accommodated, which is mounted movably parallel to the injector 142. The injection valve member 140 is surrounded by a nozzle chamber 144 and closes at its lower end injection openings 146. The nozzle chamber 144 communicates with a high-pressure bore 148 in connection, which extends substantially parallel to the injector 142, ie with an angular deviation of about 4 °. The high-pressure bore 148 communicates with a high-pressure inlet 150 and can via this from a high-pressure accumulator (common rail), which in Fig. 4 not shown, are acted upon by high pressure fuel.

The injection valve member 140 is acted upon by a nozzle spring 152 with a closing force. Furthermore, the injection valve member 140 is in communication with a control piston 154, above which a control chamber 156 is located. The control chamber 156 is surrounded by a high pressure annulus 158, which in turn communicates with the high pressure bore 148. The high-pressure annulus 158 is connected to the control chamber 156 via a control chamber throttle element 160.

The pressure in the control chamber 156 is controlled in this embodiment by a solenoid valve 162, via which a relief hole 164, which is also equipped with a throttle element, can be closed or released. Thereby, the relief hole 164 and thus the control chamber 156 is disconnected from a low pressure drain 166 and connected thereto.

If the solenoid valve 162 is opened, the pressure in the control chamber 156 decreases, and the control piston 154 and thus the injection valve member 140 move upwards and release the injection opening 146. This starts the injection process. If the solenoid valve 162 is fired, high pressure prevails in the control chamber 156, so that the injection valve member 140 is pressed into its valve seat, the injection openings 146 being closed.

The above-described system suggestions of the fluid system of the fuel injector 132 by neighbor injections thus become in the embodiment according to FIG Fig. 4 excited by the high-pressure inlet 150, which is in communication with the common rail. The pressure fluctuations propagate through the high-pressure bore 148 and thus cause, as described above, typical pressure oscillations in the nozzle chamber 144 and in the high-pressure annulus 158 with amplitudes of typically up to 300 bar.

At the in Fig. 4 illustrated embodiment of the fuel injector 132, a damping measure according to the invention is implemented to dampen these pressure fluctuations. For this purpose, a vibration damping hole 168 is provided, welene has a damping channel 170. This damping channel 170 extends essentially the same length to the high-pressure bore 148 through the injector body 134 and likewise runs essentially parallel to the injector axis 142. The angle deviations from the parallelism essentially correspond to the angular deviations of the high-pressure bores 148.

The vibration damping bore 168 connects the high pressure annulus 158 to the nozzle space 144. For this purpose, the damping passage 170 in this embodiment is connected via an annulus throttle 172 and to the nozzle space 144 via a nozzle space restrictor 174.

By the in Fig. 4 illustrated damping measure can be in a simple manner (ie only by additional implementation of the damping channel 170 and the throttles 172, 174) selectively modify the fluidic vibration behavior of the fuel injector 132. In particular, in this way the natural frequencies of the vibration system can be detuned and the sensitivity of the system to external stimulation can be significantly reduced.

Here are in the in Fig. 4 illustrated embodiment, a length of the damping channel 170 of 90 mm and a diameter of about 2 mm proved suitable. The throttles 172, 174 each have a diameter of 0.3 mm in this exemplary embodiment. The vibration damping bore 168 thus represents a total of a second fluidic high pressure system, which is arranged parallel to the high pressure bore 148. Vibration damping is essentially caused by the combination of the inertial fluid mass within the damping channel 170 and the throttles 172, 174, which is similarly describable as electrical damping in an RC resonant circuit.

In the Fig. 4 shown damping action changes the transmission behavior of the fuel injector 132 in a positive sense. In Fig. 5 For example, a transfer function for a standard injector (curve 176) is compared to a transfer function 178 of an injector having a vibration damping bore 168. The transmission function η is plotted in dB, which is the quotient of the derivative of the needle force and the derivative of the rail pressure. The application takes place in the frequency domain.

The two curves 176, 178 correspond to a fuel injector 132 according to Fig. 4 wherein curve 176 represents a fuel injector without the vibration damping bore 168 (ie, without damping passage 170, without annulus throttle 172 and without nozzle space throttle 174), curve 178, on the other hand, represents a fuel injector 132 Fig. 4 with the said elements, ie with a vibration damping bore 168.

This in Fig. 5 shown diagram, which is also referred to as Bode diagram, clearly shows two effects of the vibration damping measure: on the one hand, the natural frequencies marked as characteristic peaks in the curves 176, 178 shift through the vibration damping measure towards smaller frequencies. On the other hand, the peaks in the transfer functions are also significantly minimized, so that a total of broadband attenuation in the entire frequency range is recorded. Thus, the sensitivity of the fuel injector 132 against pressure fluctuations is significantly minimized. Particularly in the operating frequency range between 100 Hz and 4 KHz, which is particularly relevant for the wear behavior of fuel injectors, the amplitude responses due to the vibration damping bore 168 have been significantly reduced.

Thus arise through the vibration damping hole 168 according to the embodiment in Fig. 4 significant advantages over known vibration damping measures, such as in DE 103 07 871 A1 presented measure. There is overall a vibration damping recorded in the entire frequency range, so that the vibration measure, for example, is not specifically set to a specific operating point of the entire injector system or the internal combustion engine. This is a considerable advantage, as this, for example, the flexibility and applicability of the vibration damping system is extended not only with respect to different Injektortypen or types of machines, but also with respect to the efficiency of the individual operating points of the internal combustion engine. In addition, the vibration damping bore 168 is integrated directly in the injector housing 134, which saves additional measures outside of the fuel injector. The absorption damping is thus additional components or components within the fuel injector 132 itself.

In Fig. 6 Finally, results of wear simulations are shown. Shown are three different situations: the bar 180 shows the wear for a multiple injection without any attenuation measure, ie without any damping between common rail and nozzle space 144. This situation, which is also referred to as "DC" (direct connection), became arbitrary set to 100%.

On the other hand, the bar 182 denotes a fuel injector 132 with a so-called "Rail Injector Throttle" (RIT), with a diameter of 1.1 mm. This throttle element RIT, which is also known from the prior art, is arranged in the high-pressure inlet 150, in the rail outlet in front of the high-pressure line leading to the injector, and already causes a certain damping of pressure oscillations in the rail. Thus, this measure already reduces the wear of the fuel injector by about 20%.

The bar 184, on the other hand, illustrates a vibration damping measure according to the invention, for example, the in Fig. 4 shown vibration damping hole 168 and their effect. It can clearly be seen that the total wear is reduced from 100% to about 66% by the vibration damping hole 168.

The invention thus provides a way to efficiently reduce wear on fuel injectors 132 by simple additional damping measures. The damping measures can be implemented in particular in fuel injectors 132, in which the distance between the control chamber 150 and the nozzle chamber 144 is high, so that here a sufficient volume of fluid and a sufficient discharge path through the damping channel 170 can be provided.

Claims (5)

1. Fuel injector (132) for injecting fuel into a combustion chamber of an internal combustion engine, comprising an injection valve member (140), a nozzle chamber (144) which is connected to a high-pressure supply (150), and a control chamber (156) which controls a stroke movement of the injection valve member (140), wherein the pressure in the control chamber (156) can be switched by means of an actuator-controlled valve (162), and an additional vibration-damping bore (168) is provided between two points that are connected to the high-pressure supply (150), wherein the vibration-damping bore (168) comprises a damping duct (170) that extends in the injector housing (134) substantially parallel to the injector axis (142),
   wherein the vibration-damping bore (168) extends between a high-pressure annular chamber (158), which is adjacent to the control chamber (156), and the nozzle chamber (144), and
   the vibration-damping bore (168) also comprises at least one throttle element (172, 174), characterized in that
   the throttle element (172, 174) comprises a first throttle element (172) which connects the damping duct (170) and the high-pressure chamber (158), and
   a second throttle element (174) which connects the damping duct (170) and the nozzle chamber (144).
2. Fuel injector (132) according to Claim 1, wherein the throttle element (172, 174) has a cross section with a diameter of 0.1 to 0.8 mm, in particular of 0.2 to 0.5 mm and particularly preferably of 0.3 mm.
3. Fuel injector (132) according to one of the preceding claims, wherein the damping duct (170) comprises a cylindrical duct with a section of at least 1.0 mm, in particular of at least 1.5 mm and particularly preferably of 2.0 mm.
4. Fuel injector (132) according to one of the preceding claims, wherein the damping duct (170) has a length of at least 40 mm, in particular of at least 60 mm and particularly preferably of 90 mm.
5. Fuel injector (132) according to one of the preceding claims, wherein the fuel injector (132) comprises a high-pressure bore (148) which extends substantially parallel to the injector axis (142) and which is connected to the high-pressure supply (150), to the control chamber (156) and to the nozzle chamber (144), wherein the damping duct (170) extends substantially parallel to the high-pressure bore (148).

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