DE102013214589A1 - Switching valve for a fuel injector - Google Patents

Abstract

Switch valve (10) for a fuel injector (100) for internal combustion engines, comprising a valve piece (11) with a valve seat (12) and with the valve seat (12) cooperating, liftable valve closing armature (20). In the valve piece (11) has a drain hole (30) is formed with a circular cross-section serving as a throttle (31) first area with a diameter d1, serving as a diffuser (32) second area and serving as a subsequent flow geometry (33) third Includes area with a diameter d3. The second region has at least one section with a diameter d2 and d1 <d2 <d3. The third region B3 serving as the subsequent flow geometry (33) has a length l3, wherein the following applies for the ratio l3 / d3: 2 <l3 / d3 <4, preferably l3 / d3 is approximately equal to 2.5.

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.

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 in order to avoid or reduce turbulence of the fuel flow in the subsequent areas. This results in an optimized throttle length l ₁ of about 0.80 mm.

In an advantageous embodiment, 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.

Advantageously, the flow area 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

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

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

description

1 shows a switching valve according to the invention 10 That's here in a fuel injector 100 is used. The fuel injector 100 has a housing having a nozzle body 2 and one with the nozzle body 2 bolted valve body 40 includes. In the nozzle body 2 is a high-pressure room 3 formed, which is connected in operation with a not shown under high pressure fuel source, typically a common rail. In the pressure room 3 is longitudinally displaceable in a sleeve 6 guided nozzle needle 4 arranged, 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 is a valve piece 11 arranged on a shoulder in the valve body 40 is applied and into the pressure chamber 3 protrudes.

The opening and closing movement of the nozzle needle 4 is about the pressure in a control room 5 controlled. The control room 5 gets through the nozzle needle 4 , the sleeve 6 and the valve piece 11 limited. In the sleeve 6 is an inlet hole 7 formed the pressure chamber 3 with the control room 5 combines. In the valve piece 11 is a drain hole 30 formed the control room 5 switchable with a low pressure room 50 that connects in the valve body 40 is trained. valve piece 6 and sleeve 11 may also be in one piece as a large valve piece in other embodiments 6 . 11 be executed, so that the nozzle needle 4 in the valve piece 6 . 11 is led and both Zulaufborhung 7 as well as drain hole 30 in the valve piece 6 . 11 are formed.

In the valve housing 40 is a switching valve 10 arranged next to the valve piece 11 an electromagnet 42 , a valve closing anchor 20 and a valve pin 44 includes. In the valve housing 40 are the valve piece 11 and the electromagnet 42 with the interposition of a valve sleeve 41 through a clamping screw 43 tightly clamped. With the clamping screw 43 is the valve pin 44 firmly connected, on the axially displaceable valve closing armature 20 in an anchor hole formed in it 21 is guided. So is a quasi-pressure balanced switching valve 10 through the rigidly arranged valve pin 44 realized.

The valve closing anchor 20 is due to the power of a spring 45 between the valve pin 44 and valve closing anchor 20 is arranged, against one on the valve piece 11 trained valve seat 12 pressed. The diameter of the valve seat 12 and the diameter of the valve pin 44 are almost the same size. As a result, the hydraulic forces on the valve closing anchor 20 almost zero in the axial direction, the switching valve 10 is almost pressure balanced. This can be done constructively, for example, with a stepped anchorage hole 21 be implemented.

The low pressure room 50 is hydraulic with the low pressure return system of the fuel injector 100 connected.

The drain hole 30 is divided from the control room 5 to the low-pressure room 50 on three Areas: one as a throttle 31 serving first area, one as a diffuser 32 serving second area and a subsequent flow geometry 33 serving third area. The flow area of the throttle 31 is less than the flow area of the diffuser 32 , which in turn is less than the flow area of the subsequent flow geometry 33 ,

2 shows a detailed representation of the drain hole 30 in the valve piece 11 , The throttle 31 is arranged close to the control room in the first region and has a length l ₁ and a circular flow cross-section with the diameter d. ₁ Advantageous is their length-diameter ratio l ₁ / d ₁ ~ 3. At the throttle 31 closes the diffuser 32 with the length l ₂ in the second region, which in the illustrated embodiment as a double diffuser 32 is formed and a larger flow area than the throttle 31 comprising: The double diffuser 32 consists of a throttling first diffuser 32a with length l _2a and a second diffuser 32b with length l _2b , with the second diffuser 32b with the diameter d _{2b has} a larger flow area than the first diffuser 32a with the diameter d _2a . To the double diffuser 32 closes low pressure close to the space the following flow geometry 33 of the third area; it has the length l ₃ and the diameter d ₃ .

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 - ie of throttle 31 to the first diffuser 32a , from the first diffuser 32a to second diffuser 32b and second diffuser 32b to the following flow geometry 33 - are designed by chamfers, which can be rounded. Preferably, the transitions have throttle 31 to the first diffuser 32a and from the first diffuser 32a to second diffuser 32b one 45 ° bevel each and the transition from the second diffuser 32b to the following flow geometry 33 a 30 ° bevel.

The operation of the switching valve 10 is as follows: Before starting the injection process are the nozzle needle 4 and the switching valve 10 closed, no fuel flows into the combustion chamber of the internal combustion engine. Closed switching valve 10 means that the valve closing anchor 20 against the valve seat 12 is pressed and this seals. The control room 5 is under high pressure, which corresponds approximately to the high pressure of the common rail.

At the beginning of the injection process, the electromagnet 42 electrically energized and thereby exerts an attractive force on the valve closing anchor 20 out. Then the valve closing anchor leads 20 a lifting movement against the force of the spring 45 in the direction of the electromagnet 42 and thus lifts itself from the valve seat 12 from. The control room 5 is now over the drain hole 30 with the low-pressure room 50 connected. Fuel then flows out of the control room 5 in the low pressure room 50 , which is also referred to as a control of the fuel. The pressure in the control room 5 decreases, as more fuel over the drain hole 30 is controlled as over the inlet bore 7 accrues. To the same extent as the pressure in the control room 5 The resulting hydraulic force also drops on the nozzle needle 4 in the direction of the injection openings. This raises the nozzle needle 4 from a nozzle needle seat and releases the injection ports; Fuel flows into the combustion chamber of the internal combustion engine.

To complete the injection process, the electrical control of the electromagnet 42 completed. The electromagnet 42 exerts no attractive force on the valve closing anchor 20 off, and the valve closing anchor 20 is due to the spring force of the spring 45 again against the valve seat 12 pressed. About the inlet hole 7 becomes the control room 5 filled with high-pressure fuel until in the control room 5 the same pressure prevails as in the pressure chamber 3 or in the common rail. With the pressure in the control room 5 the hydraulic force on the nozzle needle also increases 4 in the direction of the injection openings and the nozzle needle 4 is pressed again against her nozzle needle seat. There is no more fuel flowing into the combustion chamber of the internal combustion engine.

Due to the high pressure and the injection course for the injection of the fuel into the combustion chamber of the internal combustion engine to the switching valve 10 high demands are placed on the fuel to be removed: on the one hand occur during a Schaltzyklusses large pressure differences within the control room 5 and in the drain hole 30 on. On the other hand, the fuel reaches during Absteuern especially in the drain hole 30 high flow rates. As the flow rate increases, the static pressure decreases. 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). With the increase in static pressure there The vapor bubbles condense abruptly over the vapor pressure, and they 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. Particularly vulnerable areas of the switching valve 10 with regard to cavitation damage are therefore the end of the throttle 31 , the diffuser 32 , the subsequent flow geometry 33 , the valve seat 12 , the valve closing anchor 20 and the valve pin 44 , So the high pressure areas of the switching valve 10 downstream of the throttle 31 ,

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

The task of the throttle 31 is it the fuel when driving out of the control room 5 to throttle and thus the movements of the nozzle needle 4 to control so that an optimized injection course shaping for the fuel injector 100 is achieved. The throttle function is essentially due to the diameter d _{1 of} the circular flow cross-section of the throttle 31 certainly. 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 larger than 3, so that the length l _{1 of} the throttle 31 0.7 ... 1.0 mm, in particular about 0.8 mm. A flow change due to migration of the damaged area from the diffuser 32 to the throttle 31 is avoided. Furthermore, there is less turbulence of the flowing fuel in the following areas, ie in the diffuser 32 , as with shorter chokes. Turbulence can carry the vapor bubbles near the wall, causing cavitation damage. A long throttle 31 concentrate in the diffuser 32 entering steam-laden Absteuerströmung more on the flow axis, ie on the symmetry axis of the drain hole 30 , Of course, the length-diameter ratio of the throttle 31 be chosen larger, but it would then require more space without improving the fuel flow in addition.

An inestimable disadvantage of the comparatively long throttle 31 is the deterioration of the μ value (water vapor diffusion resistance): As the throttle length increases, 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 associated with a higher resistance, which accordingly leads to undesirable turbulence. The task of the diffuser 32 is to lower the μ-value by optimized pressure recovery again. Advantageously, this object is achieved by the use of a double diffuser 32 Fulfills.

The diameter extensions in the double diffuser 32 have a decisive influence on the flow of fuel throughout the drain hole 30 and thus also on the μ-value. It is advantageous in the region of the double diffuser 32 a diameter extension designed so that the flow area 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 the size of A _2a . Thus results for the diameter d _{2a of} the circular flow cross-section of the first diffuser 32a : d _2a = 0.28 ... 0.42 mm, in particular 0.35 mm. And it results 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 or the double diffuser 32 with the first diffuser 32a and the second diffuser 32b a ratio of about 1 per diameter expansion has proven to be particularly favorable in terms of flow. This ratio achieves the best compromise between flow pattern, flow rate, μ-value reduction and cavitation.

This results in the first diffuser 32a a length l _2a = 0.28 ... 0.42 mm, in particular about 0.35 mm and for the second diffuser 32b a length l _2b = 0.40 ... 0.60 mm, in particular 0.50 mm.

Particularly favorable in terms of flow technology are the diameter expansions of the throttle 31 to the first diffuser 32a and from the first diffuser 32a to the second diffuser 32b widening 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 area of the inner edges of the double diffuser 32 which in turn could lead to the formation of vacuum zones and thus additional near-wall cavitation bubbles.

The to the double diffuser 32 subsequent subsequent flow geometry 33 has the task, with constant Abströmmenge between valve seat 12 and valve closing anchor 20 the static pressure level in the subsequent flow geometry 33 raise 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 valve seat 12 and valve closing anchor 20 kept away. Cavitation damage in the areas downstream of the diffuser 32 are minimized. At the same time, a long subsequent flow geometry lowers 33 the μ-value through the formation of a homogeneous flow pattern, which enhances the functionality of the fuel injector 100 improved.

By the previous use of the double diffuser 32 must the diameter extension for the subsequent flow geometry 33 no longer limited to twice the flow area. Advantageously, the diameter d _{3 of} the subsequent flow geometry 33 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 with a small stroke of the valve closing anchor 20 the required amount of diversion between valve seat 12 and valve closing anchor 20 can be dissipated.

The length-diameter ratio l ₃ / d _{3 of} the subsequent flow geometry 33 does not have to be greater than 3, choke 31 and double diffuser 32 already a large part of the high pressure in the control room 5 have broken down; in the following flow geometry 33 it comes to significantly less vapor bubble formation. 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 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 of throttle 31 to diffuser 32 and of diffuser 32 to subsequent flow geometry 33 can be positioned so that due to the screwing of the valve piece 11 existing static voltages are superimposed favorably with the caused by the high pressure loads dynamic stresses.

QUOTES INCLUDE IN THE DESCRIPTION

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Cited patent literature

• DE 19859537 [0001, 0002]

Claims (11)

Switching valve ( 10 ) for a fuel injector ( 100 ) for internal combustion engines comprising a valve piece ( 11 ) with a valve seat ( 12 ) and one with the valve seat ( 12 ) cooperating, liftable valve-closing anchors ( 20 ), wherein in the valve piece ( 11 ) a drain hole ( 30 ) is formed with a circular cross-section, one as throttle ( 31 ) serving a first area with a diameter d ₁ , one as a diffuser ( 32 ) serving second area and a subsequent flow geometry ( 33 Includes) serving the third portion having a diameter d _3, wherein the second region has at least one portion with a diameter d ₂ and it is d ₁ <d ₂ <d _3, characterized in that the (as subsequent flow geometry 33 ) third region B _{3 has} a length l ₃ , wherein for the ratio l ₃ / d ₃ : 2 <l ₃ / d ₃ <4, preferably l ₃ / d _{3 is} approximately equal to 2.5. Switching valve ( 10 ) according to claim 1, characterized in that the diameter d _{3 of} the subsequent flow geometry ( 33 ) 1.0 mm to 1.5 mm, preferably about 1.3 mm. Switching valve ( 10 ) according to claim 1 or 2, characterized in that the throttle ( 31 ) has a length l ₁ of about 0.80 mm and the diameter d _{1 of} the throttle ( 31 ) is about 0.25 mm. Switching valve ( 10 ) according to one of the preceding claims, characterized in that the diffuser ( 32 ) serving second region is designed as a double diffuser and has two cylindrical sections, with a to the throttle ( 31 ) subsequent first diffuser ( 32a ) with a diameter d _2a and a second diffuser ( 32b ) with a diameter d _2b , where d _2b > d _2a . Switching valve ( 10 ) according to claim 4, characterized in that the flow cross-section of the second diffuser ( 32b ) about twice as large as the flow area of the first diffuser ( 32a ). Switching valve ( 10 ) according to claim 4 or 5, characterized in that the first diffuser ( 32a ) has a length l _2a and has a length-to-diameter ratio l _2a / d _2a of about 1. Switching valve ( 10 ) according to one of claims 4 to 6, characterized in that the second diffuser ( 32b ) has a length l _2b and a length-to-diameter ratio l _2b / d _2b of about 1. Switching valve ( 10 ) according to one of claims 4 to 7, characterized in that the diameter of the first diffuser ( 32a ) d _{2a is} about 0.35 mm. Switching valve ( 10 ) according to one of claims 4 to 8, characterized in that the diameter of the second diffuser (32b) d _{2b is} about 0.50 mm. 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. Fuel injector ( 100 ) with a switching valve ( 10 ) according to one of the preceding claims, wherein the switching valve ( 10 ) is designed as a solenoid valve.

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