Classifications

• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
  • F02M61/00—Fuel-injectors not provided for in groups F02M39/00 - F02M57/00 or F02M67/00
  • F02M61/04—Fuel-injectors not provided for in groups F02M39/00 - F02M57/00 or F02M67/00 having valves, e.g. having a plurality of valves in series
  • F02M61/10—Other injectors with elongated valve bodies, i.e. of needle-valve type
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
  • F02M61/00—Fuel-injectors not provided for in groups F02M39/00 - F02M57/00 or F02M67/00
  • F02M61/04—Fuel-injectors not provided for in groups F02M39/00 - F02M57/00 or F02M67/00 having valves, e.g. having a plurality of valves in series
  • F02M61/08—Fuel-injectors not provided for in groups F02M39/00 - F02M57/00 or F02M67/00 having valves, e.g. having a plurality of valves in series the valves opening in direction of fuel flow
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
  • F02M—SUPPLYING COMBUSTION ENGINES IN GENERAL WITH COMBUSTIBLE MIXTURES OR CONSTITUENTS THEREOF
  • F02M61/00—Fuel-injectors not provided for in groups F02M39/00 - F02M57/00 or F02M67/00
  • F02M61/16—Details not provided for in, or of interest apart from, the apparatus of groups F02M61/02 - F02M61/14
  • F02M61/20—Closing valves mechanically, e.g. arrangements of springs or weights or permanent magnets; Damping of valve lift
  • F02M61/205—Means specially adapted for varying the spring tension or assisting the spring force to close the injection-valve, e.g. with damping of valve lift

Definitions

• the present invention relates to a fuel injector for an internal combustion engine, particularly a large two-stroke diesel engine, having an external housing for mounting in a cylinder cover and a through-going fuel passage opening out into a nozzle, and having a valve slider which is longitudinally displaceable in a slider guide and is acted upon in a direction towards its valve seat by a preloaded compression spring and in the opposite direction by the pressure of the fuel in the fuel passage, a first piston, which is axially displaceable in the housing, being located at one end of the compression spring and having a first surface which faces away from the spring and, together with a stationary component, delimits a first pressure chamber which communicates with the fuel passage through a channel, said compression spring biasing the first piston in a direction towards an extreme position with a minimum fuel volume in the first pressure chamber.
• the channel from the fuel passage to the first pressure chamber is designed as a throttle channel so that the pressure in the chamber follows the current pressure in the fuel passage with a delay.
• the pressure also increases in the pressure chamber so that the first piston is pressed towards the closing spring and increases the force with which the spring acts on the valve slider in a direction towards its seat, which increases the closing pressure of the injector.
• the closing pres ⁇ sure i.e. the pressure at which the valve slider moves towards its seat at the end of the delivery period
• the opening pressure i.e. the pressure in the fuel passage at which the valve slider is lifted off its valve seat at the beginning of a delivery period.
• the lower closing pressure is due to the fact that in the closed position of the injector, the fuel pressure acts on an effective area of the valve slider which is smaller than when the valve slider is in its open position, when the pressure also acts on the slider area below the seat surface.
• the first piston makes continuous adjustment movements during and immediately following the delivery period of the injector, which may result in a not inconsiderable wear on the guide surface of the piston with consequent increased leakage from the pressure chamber.
• the compression spring presses the first piston back to its extreme position with a minimum fuel volume in the pressure chamber, so that the opening pressure of the injector is not influenced by the hydraulic loading of the compression spring.
• the compression pressure in the engine cylinder depends on the engine load so that at full loads the pressure is substantially higher than at low loads.
• the compression pressure may be at about 120 bar, while at idling load the compression pressure is at about 40 bar.
• the pressure in the engine cylinder spreads through the nozzle holes and further along to the slider area below the seat surface, i.e. the slider section positioned at the nozzle side of the valve seat. Therefore, the current compression pressure acts on the valve slider with a force in the opening direction.
• the compression pressure rising with the engine load thus leads to a drop in the opening pressure of the known injectors at increased engine loads.
• the opening pressure may, for example, drop from 400 bar at idling load to 325 bar at full engine load. The lower opening pressure at full load does not promote the atomization of the fuel at the start of the injection period.
• the fuel pressure in the injector is determined by its opening pressure, as the fuel pumps supply such a small amount of fuel that the flow resistance in the nozzle has no influence on the fuel pressure. Contrary to this, the fuel amounts delivered from the pumps at higher engine loads are so large that the flow resistance in the nozzle becomes decisive to the fuel pressure in the injector, i.e. the fuel pressure is in this case substantially higher than the opening pressure of the injector.
• the opening pressure in the known injectors is determined by the preloading of the compression spring.
• the manufacture of the springs is subject to certain production tolerances, resulting in the fact that the fuel injectors in an internal combustion engine are not necessarily all set to exactly the same opening pres ⁇ sure.
• the variations in the opening pressures of the injectors often become more pronounced after a long period of operation of the engine, as the springs sag during operation, i.e. loose some of their spring force.
• there are time-determined changes in the opening pressures of the injectors which necessitates control and retightening of the compression springs at regular intervals to maintain satisfactory operation of the engine. This is labour-consuming and undesirable.
• the object of the invention is to provide a fuel injector which has an increased opening pressure at increasing engine loads, and which requires less maintenance.
• the fuel injector according to the invention is characterized in that the spring is in force-transmitting connection with a second piston having a second surface which faces away from the spring and constitutes the end wall in a second pressure chamber, that at movement in a direction away from said extreme position the first piston opens a flow connec ⁇ tion between the first and the second pressure chambers, that the second pressure chamber has a larger effective cross-sectional area than the first chamber, and that a restricted drain passage connects the second pressure chamber with a drain.
• the pressure grows in the fuel passage as a consequence of the flow resistance in the nozzle.
• the pressure in the first pressure chamber increases, which causes displacement of the first piston so that the flow connection between the first and the second pressure chambers is opened, and an amount of fuel flows into the second pressure chamber.
• this pressure chamber has a larger effective cross-sectional area than the first chamber, the pressure build-up in the second chamber will lead to a return of the first piston to its extreme position with a minimum fuel volume, and at the same time the amount of fuel in the second pressure chamber is confined, as the first piston interrupts the flow connection between the chambers.
• the second piston As the second piston is in force-transmitting connection with the spring, the latter will be shortened in step with the filling of the chambers, whereby the spring force increases. Apart from a small amount of fuel drained away, the fuel is kept confined in the pressure chamber, until the fuel injector again has to be actuated for renewed injection of fuel, and thus, the increased force of the spring is maintained, which imparts to the fuel injector both a higher opening pressure and a higher closing pressure.
• the fuel pressure in the first pressure chamber will generate a force on the first piston which is larger than the opposite spring force, which leads to displacement of the first piston so that the flow connection between the chambers is opened, until the pressure build-up in the second pressure chamber has generated an increased spring force slightly larger than the fuel pressure on the effective cross-sectional area of the first pressure chamber.
• the increased spring force will then again return the first piston to its extreme position with a blocked flow connection between the chambers.
• the restricted drain passage provides continuous drainage of a small amount of fuel from the second pressure chamber. This ensures that the loading of the spring is also reduced when the engine load and thus the highest pressure in the fuel passage decreases. If the load does not decrease, the amount of fuel drained away will be replaced by new fuel at the next injection period as the draining away of fuel produces a small lowering of pressure in the second pressure chamber so that the first piston can again open the flow connec ⁇ tion.
• the opening pressure of the fuel injector at a specific load in the upper load range of the engine depends on the effective cross-sectional area of the first chamber. As such an area can be manufactured at very fine tolerances, all the fuel injectors in the engine will adjust themselves with the same opening and closing pressures, as the various manufacturing-deter- mined variations in the spring characteristics of the compression springs and their different sagging during the operating period will be equalized by compression of the springs until they yield the same spring force which adjusts to the highest fuel pressure in the injectors. This adjustment takes place automatically during the operation of the engine, and this obviates a substantial part of the need for periodical manual adjustment of the injectors.
• the first piston has a seat section which blocks the flow connection between the first and the second pressure chambers by contact with a corresponding seat section on the stationary component.
• seat sections have proved very reliable in fuel injectors over many years, and they produce a well-defined closure capable of resisting large pressure differences.
• the drain passage may suitably be of such limited size that the volume drained away at full engine load during an engine cycle ranges from one half to one twentieth of the fuel volume in the second pressure chamber. If the amount drained away becomes greater than half, it will be difficult, particularly at low engine loads, to achieve the desired increase in the opening pressure, and in addition, the piston movements will become large and frequent, as the second pressure chamber will have to be refilled at each injection period, also when the engine load is constant. If the volume drained away is less than one twentieth, the opening pressure will drop unsuitably slowly at a sudden reduction of the engine load. The said drain ratios apply at full engine load.
• the effective cross-sectional area of the first pressure chamber may be smaller than the opening area of the valve slider.
• the second pressure chamber has an effective cross-sectional area which is several times larger than that of the first chamber.
• the pressure in the second pressure chamber is a corresponding number of times smaller than the pressure in the first chamber, when the force of the second piston acting on the spring and thus on the first piston equalizes the oppositely directed force of the fuel acting on the first piston.
• the large effective cross-sectional area of the second chamber thus results in closure of the second chamber at an advantageously low chamber pressure, which results in a relatively small pressure drop over the drain passage with a consequent small draining away of fuel from the second pressure chamber.
• the large area of the second chamber also yields the advantage that the chamber is filled with a large fuel volume at a certain displacement of the second piston and a corresponding compression of the spring.
• the pistons participate in the adjust ⁇ ment movements of the valve slider.
• the pistons will act as an increase of the mass of the slider, which will give the injector slower adjustment movements.
• the first piston may alternatively be formed at the opposite end of the spring. This results in the disadvantage that the flow connection between the two pressure chambers becomes elongated and relatively difficult to manufacture.
• the fuel injector is formed so that in a manner known per se the compression spring is mounted between two spring guides which are longitudinally displaceable on a central thrust member which is stationary in the housing, that the second piston is formed in the upper spring guide which is located opposite the nozzle and has a lower tubular wall which pressure-sealingly encloses the thrust member, and an upper tubular wall which has a larger internal diameter than the lower wall and pressure-sealingly encloses the first piston, and an intermediate member interconnecting the walls, its upper surface constituting the second surface, that the first piston is annular and is enclosed between the thrust member and the upper wall of the second piston and has a lower inward collar the upper surface of which constitutes the first surface, which inwardly merges into the seat section, and that the corresponding seat section of the thrust member faces downwards and is positioned between a through- going channel to the fuel passage and a lower area of reduced diameter constituting the flow connection between the two chambers
• the drain passage may be formed as an independent member, for example in the shape of a small bore through the second piston into the second chamber, but preferab ⁇ ly, the drain passage consists of the pressure-sealing 5 annular slits between the two walls of the second piston, and the first piston and the thrust member, respectively, these annular slits being difficult to make completely pressure-tight as it is.
• the amount of fuel drained away will at the same time lubricate the 10 surfaces sliding against each other.
• Fig. 1 shows a partial, longitudinal section 15 through a fuel injector according to the invention
• Fig. 2 on a larger scale, shows a portion of Fig. 1 depicting the compression spring with associated members
• Fig. 3 is a diagram of the correlation between the
• Fig. 1 shows a fuel injector generally designated
• the housing is elongated and at its
• source of periodic supply of highly pressurized fuel is connected through a pressure conduit to a fuel inlet 5 at the top of the injector, from where a fuel passage
• the fuel passage may have a valve which opens for circulation of preheated fuel in the injector between injection periods.
• the fuel passage passes through a stationary component in form of a thrust member 9 which is in upward contact with a member 10 which is station ⁇ ary in the housing, and in downward contact with an intermediate member 11 which is pressed firmly against a slider guide 12 stationary in the housing.
• a valve slider 13 is received so as to be longitudinally displaceable in a central guide bore 12' in the slider guide, and by its one end encloses a downwardly project ⁇ ing cylindrical section 11' on the intermediate member.
• the guide bore centres the slider so that an annular conical seat surface 14 located at the lower end of the valve slider and formed as a needle, is coaxial with a corresponding valve seat on the slider guide 12.
• the tip of the needle projects into the central cavity of the nozzle and is here exposed to the engine cylinder pressure which spreads through the nozzle holes into the cavity and acts on the valve slider with a force in the opening direction.
• the lower end of the valve slider and the slider guide 12 delimit a pressure chamber 15 which communicates with the fuel passage 6 via oblique bores 16.
• the downward, annular end surface of the valve slider, which is inwardly delimited by the needle, is influenced by the fuel pressure in the chamber 15, as the fuel pressure acts on the valve slider with a force in the opening direction.
• the opening area of the valve slider is substantially determined by the difference in diameters between the external diameter of the cylindrical section 11' and the internal diameter of the guide bore 12' .
• the valve slider is also loaded in the closing direction, i.e. in a downward direction towards the valve seat, by means of a compression spring 17, the upper end of which is in contact with an upper spring guide 18 mounted displaceably on the thrust member 9, and the lower end of which is supported via a lower spring guide 19 also displaceably guided on the thrust member 9, by a slotted thrust bushing, the lower end surface of which is in contact with an upward collar on the valve slider 13.
• the spring force is thus trans ⁇ mitted via the spring guide 19 and the thrust bushing 20 to the valve slider 13.
• An annular first piston 21 is mounted axially displaceably around the upper section of the thrust member 9.
• the internal diameter of the sliding surface 22 (Fig.
• the guide surface 23 terminates downwards in an cylindrical recess worked into the thrust member and communicating through a channel 24 with the central fuel passage 6 in the thrust member.
• the recess merges downwards into a cylindrical section 25 of a smaller external diameter than the guide surface 23.
• the thrust member has an annular conical downward seat section 26.
• the first piston 21 has a lower inward collar 27 having a conical upward seat section 26' which may pressure-sealingly abut the seat section 26.
• the first piston and the thrust member delimit a first pressure chamber 28 with an effective cross-sectional area determined by the difference in diameters between the section 25 and the guide surface 23.
• the effective cross-sectional area is positioned on the upper surface of the collar 27, i.e. on a first surface facing away from the spring so that the fuel pressure supplied through the channel 24 in the first pressure chamber acts on the first piston with a downward force.
• a second piston 29 is formed integrally with the upper spring guide 18 and comprises an annular inter- mediate member 30 which supports a lower tubular wall 31 and an upper tubular wall 32.
• the inner side of the lower wall 31 is pressure-sealingly and longitudinally displaceably in contact with a cylindrical second guide surface 33 of the thrust member 9, and the inner side of the upper wall 32 is in pressure-sealing and axially displaceable contact with the outer side of the first piston 21.
• the first and the second pistons together with the thrust member 9 delimit a second pressure chamber 34 of an effective cross-sectional area deter- mined by the difference in diameters between the cylindrical inner sides of the lower wall 31 and the upper wall 32.
• a cylindrical recess formed in the outer side of the thrust member and positioned immediately below the seat section 26 forms a flow connection 35 between the two pressure chambers, when the first piston has been moved away from the seat section 26.
• the bottom side of the collar 27 may have one or more projections or an annular projection with cuts keeping the part of the second pressure chamber 34 positioned radially in alignment with the projection in flow communication with the flow connection 35, when the projection abuts the upper side of the intermediate member 30.
• the projection may be annular, and the inner side of the lower wall 31 may have a smaller diameter than the inner side of the collar 27 so that the part of the second pressure chamber positioned closest to the flow connec ⁇ tion 35 has an upward effective cross-sectional area which is open to the flow connection 35 when the projection on the collar 27 abuts the upper side of the intermediate member 30 and blocks the connection to the remaining part of the pressure chamber 34.
• this effective area will cause the second piston 29 to be moved away from the first piston 21 with simultaneous uncovering of the full effective cross-sectional area of the second pressure chamber.
• the second pressure chamber 34 is continuously in contact with a restricted drain passage consisting of a pressure-sealing annular slit between the inner side of the upper wall 32 and the cylindrical outer side of the first piston, and of the pressure-sealing annular slit between the inner side of the lower wall 31 and the guide surface 33 on the thrust member.
• the pressure in the fuel passage 6 rises during each injection period to a highest pressure which at low loads corresponds to the opening pressure of the injector and at higher loads is determined by the flow resistance in the nozzle holes.
• the maximum pressure in the fuel passage thus rises at increasing engine loads.
• the pressure in the fuel passage 6 spreads through the channel 24 to the first pressure chamber 28, and when the pressure here reaches a level where the downward force on the first piston overcomes the preset spring force, the first piston is moved towards the spring, which is compressed between the spring guides 18 and 19, and at the same time the fuel flows via the flow connection 35 into the second pressure chamber where the pressure is built up to a level which returns the first piston 21 into contact with the seat section 26, while the second piston 29 remains in the position in which the spring is given the extra loading. If the pressure in the fuel channel 6 at subsequent injection periods increases to a higher level, the piston movements are repeated so that the spring 17 is given a loading which depends in a linear manner on the maximum pressure in the fuel passage 6.
• the fuel injector according to the invention achieves a spring force and thus an opening pressure which increases with increasing engine loads, as shown in Fig. 3. This renders it possible to lower the opening pressure at low engine loads, as the injector automatically generates the high opening pressure required at full load.
• the compression spring can be pre-manufactured with a preloading yielding an opening pressure at low loads of about 200 bar, which promotes stable engine operation at partial loads, and at the same time the opening pressure at full loads is higher than in the known injectors, which promotes good atomization of the fuel at the start of the injection period.
• the fuel passage 6 may in a well-known manner comprise a number of channels which extend in a stationary intermediate member along the outer side of the spring and which open out in the pressure chamber 15 via oblique channels in the slider guide.
• the opening area of the valve slider is determined by the downward annular end surface surrounding the needle.

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• Engineering & Computer Science (AREA)
• Chemical & Material Sciences (AREA)
• Combustion & Propulsion (AREA)
• Mechanical Engineering (AREA)
• General Engineering & Computer Science (AREA)
• Fuel-Injection Apparatus (AREA)

Priority Applications (9)

Applications Claiming Priority (2)

Publications (1)

Family

ID=8099107

Family Applications (1)

Country Status (14)

Cited By (1)

Citations (3)

• 1994
  • 1994-08-09 DK DK092794A patent/DK171216B1/da not_active IP Right Cessation
• 1995
  • 1995-07-11 CN CN95194557A patent/CN1060250C/zh not_active Expired - Fee Related
  • 1995-07-11 AU AU31612/95A patent/AU3161295A/en not_active Abandoned
  • 1995-07-11 WO PCT/DK1995/000300 patent/WO1996005425A1/en active IP Right Grant
  • 1995-07-11 EP EP95927656A patent/EP0733163B1/en not_active Expired - Lifetime
  • 1995-07-11 KR KR1019960706534A patent/KR970703488A/ko not_active IP Right Cessation
  • 1995-07-11 ES ES95927656T patent/ES2101609T3/es not_active Expired - Lifetime
  • 1995-07-11 RU RU96118505A patent/RU2126095C1/ru active
  • 1995-07-11 JP JP08506931A patent/JP3090473B2/ja not_active Expired - Fee Related
  • 1995-07-11 PL PL95315571A patent/PL176852B1/pl unknown
  • 1995-07-11 DE DE69500259T patent/DE69500259T2/de not_active Expired - Fee Related
  • 1995-08-09 HR HR950445A patent/HRP950445B1/xx not_active IP Right Cessation
• 1996
  • 1996-07-22 FI FI962931A patent/FI107470B/fi active
• 1997
  • 1997-02-04 NO NO970494A patent/NO306793B1/no not_active IP Right Cessation

Patent Citations (3)

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