EP3025048B1 - Control valve for a fuel injector - Google Patents

Description

State of the art

Fuel injectors, as they are preferably used for fuel injection directly into the combustion chamber of an internal combustion engine, are known from the prior art, such as from DE 198 59 537 , In injection systems which operate according to the so-called common rail principle, compressed fuel is made available in a rail by means of a high-pressure pump and injected by means of the fuel injectors into the respective combustion chambers of an internal combustion engine. The injection is controlled by means of a nozzle needle arranged in the fuel injector, which performs a longitudinal movement and thereby opens and closes one or more injection openings. The nozzle needle is moved by hydraulic forces acting on the nozzle needle due to a pressure in a control chamber. By changing the pressure in the control chamber and thus the closing force on the nozzle needle, the longitudinal movement of the nozzle needle can be controlled specifically. For the opening movement of the nozzle needle, the pressure in the control chamber is lowered by means of a switching valve by a hydraulic connection between the control chamber and a low-pressure chamber is opened, wherein the hydraulic connection is designed as a throttle.

From the DE 198 59 537 It is also known that a particularly advantageous injection rate shaping can be achieved by a combination of throttle and diffuser as a hydraulic connection. Due to the ever-increasing demands regarding the high pressure of the fuel to be injected, the high pressure applied in the control chamber and thus also the high pressure to be removed via the throttle increases. This increases the risk of cavitation damage in throttle and subsequent flow geometry due to increasing vapor bubble formation in the discharged fuel.

Another example is from the DE 10 2007 004553 A1 known.

The present invention has for its object to provide a switching valve with a hydraulic connection for Absteuern the fuel from the control chamber in the low-pressure chamber, in which arise as few cavitation damage within the hydraulic connection.

Disclosure of the invention

The switching valve according to the invention with the features of claim 1 achieved even at very high pressures on the one hand a particularly advantageous injection course shaping and on the other hand, the occurrence of cavitation damage is minimized.

For this purpose, the switching valve of a fuel injector for internal combustion engines comprises a valve piece with a valve seat and a cooperating with the valve seat, liftable valve-closing armature, wherein in the valve piece a drain hole is formed with a circular cross-section. The drainage bore comprises a first region serving as a throttle with a diameter d ₁ , a second region serving as a diffuser, and a third region having a diameter d ₃ serving as a subsequent flow geometry. The second region has at least one section with a diameter d ₂ and d ₁ <d ₂ <d ₃ . The third region B ₃ serving as the subsequent flow geometry has a length l ₃ , wherein the following applies to the ratio l ₃ / d ₃ : 2 <l ₃ / d ₃ <4, preferably l ₃ / d _{3 is} approximately equal to 2.5.

Advantageously, the subsequent flow geometry has a diameter d ₃ of 1.0 mm to 1.5 mm, preferably about 1.3 mm. The diameter d ₃ is only slightly smaller than the diameter of the valve seat, so that the areas near the wall are as far away as possible from the flow axis and the imploding of the cavitation bubbles leads to no or only minimal cavitation damage.

Advantageously, the throttle has a length l ₁ of about 0.80 mm and a diameter d ₁ of about 0.25 mm. The throttle diameter d ₁ of about 0.25 mm leads to a particularly favorable injection rate shaping, especially during the opening stroke of the nozzle needle. Preferably, the length-diameter ratio l ₁ / d _{1 of} the throttle is greater than 3, to turbulence of the To avoid or reduce fuel flow in the following areas. This results in an optimized throttle length l ₁ of about 0.80 mm. According to the invention, the diffuser has a second section with the diameter d _2b . As a result, the diffuser is designed as a double diffuser, with a first diffuser, which adjoins the throttle, having a diameter d ₂ or d _2a and a second diffuser having a diameter d _2b , where d _{2b is} > d _2a . Through the use of a double diffuser, the flow throughout the drain hole is made advantageous: turbulence is reduced and the imploding of the vapor bubbles in near-wall "dead water areas" - ie areas with very low flow velocity, usually at inner edges - avoided as far as possible. According to the flow cross-section of the second diffuser is about twice as large as the flow area of the first diffuser. As a result, the unfavorably high μ value (water vapor diffusion resistance) due to the throttling is lowered again.

In an advantageous embodiment, the first diffuser has a length l _2a and has a length-diameter ratio l _2a / d _2a of about 1. This ratio achieves the best compromise flow pattern, flow rate, reduction of μ value and occurrence of cavitation.

Similarly, the second diffuser advantageously has a length l _2b and has a length-diameter ratio l _2b / d _2b of about 1.

Preferably, the first diffuser has a diameter d _2a of about 0.35 mm. In the area of the double diffuser, it is particularly favorable for an enlargement of the diameter if the flow cross section per stage approximately doubles. Thus, the fuel flow with as little as possible turbulence is performed while optimizing pressure recovery. With a throttle diameter d ₁ of about 0.25 mm, this results in about 0.35 mm for the diameter d _{2a of} the first diffuser.

Similarly, the diameter of the second diffuser d _{2b is} about 0.50 mm.

Preferably, the switching valve is designed as a substantially pressure-balanced valve. As a result, high switching speeds can be achieved and the switching valve is thus very well suited for multiple injections.

Advantageously, the switching valve is designed as a solenoid valve. The structure of the switching valve according to the invention, especially the proximity of drain throttle to valve seat, promotes electromagnetic excitation of the valve closing armature. A pressure boost with an additional throttle point can be omitted.

drawings

Fig.1

shows a switching valve according to the invention in longitudinal section, wherein only the essential areas are shown.

Fig.2

shows an enlarged section of the Fig.1 with the valve piece of the switching valve, wherein a drain hole formed in the valve piece is shown in more detail.

description

Fig.1 shows a switching valve 10 according to the invention , which is used here in a fuel injector 100 . The fuel injector 100 has a housing which comprises a nozzle body 2 and a valve housing 40 screwed to the nozzle body 2 . In the nozzle body 2 , a high-pressure pressure chamber 3 is formed, which is connected in operation with a not shown under high pressure fuel source, typically a common rail. In the pressure chamber 3 , a guided in a sleeve 6 nozzle needle 4 is arranged longitudinally displaceable, which serves for opening and closing of injection openings, not shown, which open into a combustion chamber, not shown, of an internal combustion engine.

In the valve housing 40 , a valve piece 11 is arranged, which bears against a shoulder in the valve housing 40 and projects into the pressure chamber 3 .

The opening and closing movement of the nozzle needle 4 is controlled by the pressure in a control chamber 5 . The control chamber 5 is limited by the nozzle needle 4, the sleeve 6 and the valve piece 11 . In the sleeve 6 , an inlet bore 7 is formed, which connects the pressure chamber 3 with the control chamber 5 . In the valve piece 11 , a drain hole 30 is formed, which connects the control chamber 5 switchable with a low-pressure chamber 50 which is formed in the valve housing 40 . Valve piece 6 and sleeve 11 can be made in one piece as a large valve piece 6, 11 in other embodiments, so that the nozzle needle 4 in the valve piece 6, 11 is guided and both Zulaufborhung 7 and 30 drain hole in the valve piece 6, 11 are formed.

In the valve housing 40, a switching valve 10 is arranged which includes a solenoid adjacent the valve element 11 42, a valve-closing armature 20 and a valve pin 44th In the valve housing 40 , the valve member 11 and the solenoid 42 with the interposition of a valve sleeve 41 by a clamping screw 43 are firmly clamped. With the clamping screw 43 of the valve pin 44 is fixedly connected, is guided on the axially displaceable valve closing armature 20 in an anchor bore 21 formed therein. Thus, a quasi-pressure-balanced switching valve 10 is realized by the rigidly arranged valve pin 44 .

The valve-closing armature 20 is urged by the force of a spring 45 which is disposed between valve pin 44 and valve-closing armature 20, against a piece 11 formed on the valve seat 12th The diameter of the valve seat 12 and the diameter of the valve pin 44 are almost equal. As a result, the hydraulic forces on the valve-closing armature 20 in the axial direction are virtually zero, the switching valve 10 is quasi pressure-balanced. Constructively, this can be implemented, for example, by a stepped anchorage bore 21 .

The low-pressure chamber 50 is hydraulically connected to the low-pressure return system of the fuel injector 100 .

The drainage bore 30 is subdivided from the control chamber 5 to the low-pressure chamber 50 into three regions: a first region serving as a restrictor 31 , a second region serving as a diffuser 32 and a third region serving as a subsequent flow geometry 33 . The flow cross section of the throttle 31 is less than the flow cross section of the diffuser 32, which in turn is less than the flow area of the subsequent flow geometry 33.

Fig.2 shows a detailed representation of the discharge bore 30 in the valve piece 11. The throttle 31 is arranged steuerraumnah in the first region and has a length l ₁ and a circular flow cross-section of diameter _d1. Its length-diameter ratio l ₁ / d ₁ ~ 3 is advantageous. The diffuser 32 with the length l ₂ adjoins the throttle 31 in the second region, which is designed as a double diffuser 32 and has a larger flow cross-section than the throttle 31 The double diffuser 32 is composed of a throttle-proximate first diffuser 32a of length l _2a and a second diffuser 32b of length l _2b , the second diffuser 32b of diameter d _2b having a larger flow area than the first diffuser 32a of diameter d _2a . The subsequent flow geometry 33 of the third region adjoins the double diffuser 32 at low pressure space; it has the length l ₃ and the diameter d ₃ .

• d ₁< d ₂ < d ₃ für eine Ablaufbohrung 30 mit einfachem Diffusor 32
• d ₁< d _2a <d _2b< d₃ für eine Ablaufbohrung 30 mit Doppeldiffusor 32a & 32b

All three areas of the drain hole 30 are characterized by circular flow cross-sections. The following applies: The flow cross-sections always extend from the control room side to the low-pressure room side:

• d ₁ < d ₂ < d ₃ for a drain hole 30 with a simple diffuser 32
• d ₁ < d _2a <d _2b < d ₃ for a drain hole 30 with double diffusers 32a & 32b

The transitions between the different flow cross-sections - that is from throttle 31 to first diffuser 32a, from first diffuser 32a to second diffuser 32b and from second diffuser 32b to the subsequent flow geometry 33 - are configured by chamfers, which may be rounded. Preferably, the transitions from throttle 31 to first diffuser 32a and from first diffuser 32a to second diffuser 32b each have a 45 ° phase and the transition from second diffuser 32b to the subsequent flow geometry 33 has a 30 ° phase.

The operation of the switching valve 10 is as follows: Before the start of the injection process, the nozzle needle 4 and the switching valve 10 are closed, there is no fuel flowing into the combustion chamber of the internal combustion engine. closed Switching valve 10 means that the valve-closing armature 20 is pressed against the valve seat 12 and seals it. The control chamber 5 is at high pressure, which corresponds approximately to the high pressure of the common rail.

At the beginning of the injection process, the solenoid 42 is electrically energized and thereby exerts an attractive force on the valve-closing armature 20 . Thereafter, the valve closing armature 20 performs a lifting movement against the force of the spring 45 in the direction of the electromagnet 42 and thus stands out from the valve seat 12 . The control chamber 5 is now connected via the drain hole 30 to the low-pressure chamber 50 . Fuel then flows from the control chamber 5 in the low-pressure chamber 50 , which is also referred to as Absteuerung of the fuel. The pressure in the control chamber 5 decreases because more fuel is diverted via the drain hole 30 than flows through the inlet bore 7 . To the same extent as the pressure in the control chamber 5 , the resulting hydraulic force on the nozzle needle 4 decreases in the direction of the injection openings. As a result, the nozzle needle 4 lifts off from a nozzle needle seat and releases the injection openings; Fuel flows into the combustion chamber of the internal combustion engine.

To end the injection process, the electrical control of the electromagnet 42 is stopped. The solenoid 42 no longer exerts any attractive force on the valve-closing armature 20 , and the valve-closing armature 20 is pressed against the valve seat 12 again by the spring force of the spring 45 . About the inlet bore 7 of the control chamber 5 is filled with high-pressure fuel until the same pressure prevails in the control chamber 5 as in the pressure chamber 3 and the common rail. With the pressure in the control chamber 5 , the hydraulic force on the nozzle needle 4 increases in the direction of the injection openings and the nozzle needle 4 is pressed against its nozzle needle seat again. There is no more fuel flowing into the combustion chamber of the internal combustion engine.

Due to the high pressure and the injection curve for the injection of the fuel into the combustion chamber of the internal combustion engine to the switching valve 10 high demands are made with respect to the fuel to be controlled: on the one hand occur during a Schaltzyklusses large pressure differences within the control chamber 5 and in the drain hole 30 . On the other hand, the fuel reaches high flow velocities during the shutdown especially in the drain hole 30 . With increasing flow speed of the sinks static pressure. If this falls below the vapor pressure of the fuel, then vapor bubbles form. The vapor bubbles are conveyed with the flow in areas of higher static pressure (ie lower flow rate or larger flow area). As the static pressure increases above the vapor pressure, the vapor bubbles suddenly condense and implode. This process is called cavitation. For the surrounding flow geometry, the reflected implosion bubbles, which may have pressure peaks up to 40 kbar, are detrimental if they occur close to the wall; it comes to material erosion, this is called cavitation erosion or cavitation damage. Especially vulnerable areas of the switching valve 10 with respect. Cavitation damages are thus the end portion of the throttle 31, the diffuser 32, the subsequent flow geometry 33, the valve seat 12, the valve-closing armature 20 and the valve pin 44, so the highly pressurized areas of the switching valve 10 downstream of the throttle 31.

In order to prevent or reduce cavitation damage, two measures are taken with regard to the design of the drain hole 30 : on the one hand, the near-wall formation and management of vapor bubbles is reduced by the special design of the throttle 31 and the diffuser 32 , on the other hand, the imploding of the vapor bubbles distributed by the use of a subsequent flow geometry 33 with a large volume on a large, as far away from the wall area.

The task of the throttle 31 is to throttle the fuel when it is being scavenged out of the control chamber 5 and thus to control the movements of the nozzle needle 4 in such a way that an optimized injection curve shaping for the fuel injector 100 is achieved. The throttle function is essentially determined by the diameter d _{1 of} the circular flow cross-section of the throttle 31 . Hydraulic interpretations of the injection progression molding show that with a throttle diameter d ₁ of 0.2... 0.3 mm, in particular approx. 0.25 mm, the most favorable injection profile formations are achieved.

The length-diameter ratio of the throttle 31 is ideally slightly greater than 3, so that the length l _{1 of} the throttle 31 is 0.7 ... 1.0 mm, in particular about 0.8 mm. A flow change by moving the damage area from the diffuser 32 to the throttle 31 is thereby avoided. It continues to happen less strong turbulence of the fuel flowing through in the following areas, ie in the diffuser 32, than in the case of shorter throttles. Turbulence can carry the vapor bubbles near the wall, causing cavitation damage. A long throttle 31 concentrates the steam-laden bleed flow entering the diffuser 32 more on the flow axis, ie on the symmetry axis of the drainage bore 30. Of course, the length-diameter ratio of the throttle 31 can also be selected larger, but would then require more space, without additionally improving the fuel flow.

An unavoidable disadvantage of the comparatively long throttle 31 is the deterioration of the μ value (water vapor diffusion resistance): With increasing throttle length, the μ value increases, which in turn results in a high return back pressure. As a result, the flow or the "diffusion" of the vapor bubbles in the subsequent flow geometry 33 is associated with a higher resistance, which accordingly leads to undesirable turbulence. The task of the diffuser 32 is to lower the μ value again by optimizing pressure recovery. Advantageously, this object is achieved by the use of a double diffuser 32 .

The diameter expansions in the double diffuser 32 have a decisive influence on the fuel flow in the entire drain hole 30 and thus also on the μ value. In the region of the double diffuser 32 , an enlargement of the diameter is designed so that the flow cross section per stage approximately doubles. That is, the flow area A _{2a of} the first diffuser 32a is about twice as large as the flow area A _{1 of} the throttle 31 and the flow area A _{2b of} the second diffuser 32b is about twice as large as A _2a . Thus, for the diameter d _{2a of} the circular flow cross section of the first diffuser 32a, d _2a = 0.28 to 0.42 mm, in particular 0.35 mm. And it follows for the diameter d _{2b of} the circular flow cross-section of the second diffuser 32b : d _2b = 0.40 ... 0.60 mm, in particular about 0.50 mm.

For the length-diameter ratio of the diffuser 32 and the double diffuser 32 with the first diffuser 32a and the second diffuser 32b , a ratio of about 1 per diameter enlargement is particularly fluid favorably exposed. This ratio achieves the best compromise between flow pattern, flow rate, μ-value reduction and cavitation.

This results in a length l _2a = 0.28 ... 0.42 mm, in particular about 0.35 mm for the first diffuser 32a and a length l _2b = 0.40 ... 0 for the second diffuser 32b , 60 mm, in particular 0.50 mm.

Particularly favorable in terms of flow technology, the diameter expansions of restrictor 31 to first diffuser 32a and from first diffuser 32a to second diffuser 32b extend conically at 45 °, ie with a 45 ° bevel, which can be rounded at its edges. In sharp-edged, ie 90 ° transitions, so-called dead water areas would develop in the region of the inner edges of the double diffuser 32 , which in turn could lead to the formation of negative pressure zones and thus additional cavitation bubbles close to the wall.

The subsequent to the double diffuser 32 subsequent flow geometry 33 has the function of raising at a constant Abströmmenge between valve seat 12 and valve-closing armature 20, the static pressure level in the subsequent flow geometry 33 in order to achieve a reduced vapor formation. This is achieved over a comparatively large volume. The remaining vapor bubble formation and above all implosion of the vapor bubbles is also distributed to a larger volume and kept away from valve seat 12 and valve closing anchor 20 . Cavitation damage in the areas downstream of the diffuser 32 is thus minimized. At the same time, a long subsequent flow geometry 33 lowers the μ value by forming a homogeneous flow pattern, which improves the functionality of the fuel injector 100 .

Due to the previous use of the double diffuser 32 , the diameter extension for the subsequent flow geometry 33 no longer has to be limited to twice the flow cross section. Advantageously, the diameter d _{3 of} the subsequent flow geometry 33 is only slightly smaller than the diameter of the valve seat 12 and is in the range 1.0 ... 1.5 mm, preferably 1.3 mm. The comparatively large diameter of the valve seat 12 has the advantage that even at low lift of the valve closing armature 20 the required Absteuermenge between valve seat 12 and valve closing armature 20 can be removed.

The length-diameter ratio l ₃ / d _{3 of} the subsequent flow geometry 33 need not be greater than 3, since throttle 31 and double diffuser 32 have already degraded a large part of the high pressure in the control chamber 5 ; in the following flow geometry 33 , significantly less vapor bubble formation occurs. However, the length l _{3 must} be chosen so that there is a sufficient decrease in the μ value. Typically, l ₃ / d ₃ = 2 ... 4, ideally l ₃ / d ₃ = about 2.5 is chosen. This results in the following flow geometry 33 has a length l ₃ = 2.0 ... 6.0 mm, in particular about 3.25 mm.

A long subsequent flow geometry 33 also has the advantage that the stress critical areas of the drain hole 30 , namely the diameter transitions from throttle 31 to diffuser 32 and diffuser 32 to subsequent flow geometry 33 can be positioned so that the existing due to the screwing of the valve piece 11 static Strains are superimposed favorably with the dynamic stresses caused by the high-pressure loads.

Claims (9)

1. Switching valve (10) for a fuel injector (100) for internal combustion engines, comprising a valve piece (11) with a valve seat (12) and comprising a valve closing armature (20) which interacts with the valve seat (12) and which can perform stroke movements, wherein an outflow bore (30) is formed in the valve piece (11), which outflow bore has a circular cross section and which outflow bore has a first region, which serves as throttle (31) and which has a diameter d₁, a second region, which serves as diffuser (32), and a third region, which serves as a downstream flow geometry (33) and which has a diameter d₃, wherein the second region has at least one section with a diameter d₂, and d₁ < d₂ < d₃ applies, wherein the third region B₃ which serves as downstream flow geometry (33) has a length l₃, wherein, for the ratio l₃/d₃, 2 < l₃/d₃ < 4 applies, wherein the second region which serves as diffuser (32) is formed as a double diffuser and has two cylindrical sections, with a first diffuser (32a), which adjoins the throttle (31) and which has a diameter d_2a, and with a second diffuser (32b), which has a diameter d_2b, wherein d_2b > d_2a applies, characterized in that the throughflow cross section of the second diffuser (32b) is approximately twice as large as the throughflow cross section of the first diffuser (32a).
2. Switching valve (10) according to Claim 1,
   characterized in that the diameter d₃ of the downstream flow geometry (33) amounts to 1.0 mm to 1.5 mm, preferably approximately 1.3 mm.
3. Switching valve (10) according to Claim 1 or 2,
   characterized in that the throttle (31) has a length l₁ of approximately 0.80 mm, and the diameter d₁ of the throttle (31) amounts to approximately 0.25 mm.
4. Switching valve (10) according to one of the preceding claims,
   characterized in that the first diffuser (32a) has a length l_2a and a length-diameter ratio l_2a/d_2a of approximately 1.
5. Switching valve (10) according to one of the preceding claims,
   characterized in that the second diffuser (32b) has a length l_2b and a length-diameter ratio l_2b/d_2b of approximately 1.
6. Switching valve (10) according to one of the preceding claims,
   characterized in that the diameter of the first diffuser (32a) d_2a is approximately 0.35 mm.
7. Switching valve (10) according to one of the preceding claims,
   characterized in that the diameter of the second diffuser (32a) d_2b is approximately 0.50 mm.
8. Switching valve (10) according to one of the preceding claims,
   characterized in that the switching valve (10) is formed as a substantially pressure-balanced valve.
9. Fuel injector (100) having a switching valve (10) according to one of the preceding claims, wherein the switching valve (10) is formed as a solenoid valve.

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