Fuel Injection Apparatus For Internal Combustion Engines

Abstract

There is described a pump-and-nozzle assembly which forms part of a fuel injection apparatus serving an internal combustion engine and which includes a reciprocating pump piston performing its injection strokes by virtue of hydraulic pressure intermittently applied thereto in a controlled manner. Between two injection strokes the pump work chamber is charged with pressurized fuel through a supply valve and a throttle which causes the charging period to be substantially longer than the injection period. The supply valve is biased closed by a spring pressure which, together with the maximum charging pressure prevailing in the pump work chamber, is at most equal to the fuel supply pressure which, in turn, is greater than the opening pressure of the fuel injection valve disposed downstream of and in communication with the pump work chamber.

Description

BACKGROUND OF THE INVENTION

This invention relates to a fuel injection apparatus, particularly for use with externally ignited internal combustion engines. The fuel injection apparatus is of the type which has a hydraulically driven pump piston, one radial face of which bounds the pump work chamber, while the other radial face of which bounds a servo pressure chamber. The latter, for initiating the pressure stroke of the pump, is placed under pressure by fuel which, in turn, is pressurized to the fuel supply pressure by a pressure source and admitted by a conduit to the servo pressure chamber. Between the pressure source and the pump work chamber there is disposed a supply valve. The fuel injection apparatus is further provided with a solenoid valve which controls the fuel flow from the pressure source to the servo pressure chamber and from the servo pressure chamber to a discharge conduit. The fuel injection apparatus is further of the type that has a spring-biased fuel injection valve, the opening pressure of which is greater than the charging pressure that prevails in the pump work chamber during and at the termination of the charging stroke.

In a known fuel injection apparatus of the aforeoutlined type, as disclosed, for example, in U.S. Pat. No. 2,598,528, a solenoid control valve controls a servo piston which has a diameter that is larger than that of the pump piston. The servo piston and the pump piston are connected to one another; their stroke together determines the injected fuel quantities. This fuel injection apparatus, by virtue of pressure transformation, is adapted for very high injection pressures (350-1,050 kg/cm.sup.2) as it is required in diesel engines operating on direct injection. Such a fuel injection apparatus, however, is very expensive and, because of plunger control, it is too slow for high rpm engines.

For engines which require lower injection pressures, particularly engines operating on fuel mixture and ignited by spark plugs (Otto engines) and also, diesel engines with injection pressures up to approximately only 200 kg/cm.sup.2, there are required less expensive, simpler and fast-operating fuel injection apparatuses.

Further, there is also known a fuel injection apparatus, as disclosed, for example, in German Pat. No. 535,494, which operates without an additional servo piston and in which both effective radial faces of the pump piston are of identical area. The injection pressure is, however, generated by means of a spring exerting a force on the pump piston, and by means of the supply pressure. The charging period and thus the injection quantities are determined by a counterpressure which prevails in the servo chamber of the pump piston and which is variable by a regulated throttle member. The moments of the injection start and the beginning of the charging stroke are determined by a mechanically driven control plunger valve. This fuel injection apparatus has the disadvantage that because of the springs, the long fuel lines necessitated by the system and because of the mechanical control and the pressure control of the counter-pressure in the servo pressure chamber, a sufficiently rapid, accurate operation required for engines of current design cannot be achieved. Furthermore, this fuel injection apparatus is rpm-dependent, since the charging period controlled by the cam shaft varies in proportion to the rpm, whereas the throttle effect of the throttle member does not. Since both effects are superimposed, a constant injected fuel quantity cannot be achieved in case of a rapidly changing rpm.

In Otto engines operating on intermittent fuel injection, there are known electrically controlled fuel injection apparatuses with solenoid fuel injection valves (such as disclosed, for example, in German Pat. No. 1,100,377), in which the injected fuel quantities are determined by controlling the "open" period of the electromagnetically operated fuel injection valve. In solenoid injection valves of this type the injected fuel quantities are in a very substantial extent dependent upon the outlet area of the nozzle opening; thus, extremely narrow tolerances in the manufacture of the nozzle bores are required. Further, it is to be considerated a disadvantage that a change in the outlet area of the nozzle opening caused by gumming or soiling of the nozzle openings leads to undesired changes in the injected fuel quantities. It is a further disadvantage of an electromagnetic fuel injection valve of this type that in case the valve seat does not seal properly or if the valve needle or the magnet is jammed, the valve injects fuel continuously which may lead to engine failure. It is also a disadvantage that in case of small injected fuel quantities, the atomization of the injected fuel is not satisfactory because the duration of injection for such small quantities is extremely short. Consequently, the valve needle is not capable of executing a full stroke and because of the reduced stroke, the fuel is throttled through the narrow gap prevailing in such a case at the valve seat. Such a throttling results in an unsatisfactory atomization.

These last-named disadvantages are not present in fuel injection apparatuses of the type discussed earlier, but their use particularly in Otto engines is precluded because of the disadvantages set forth in connection therewith.

OBJECT, SUMMARY AND ADVANTAGES OF THE INVENTION

It is an object of the invention to provide a fuel injection apparatus which is particularly adapted for use in Otto engines operating on injected fuel and which, in addition to the advantages of the afore-outlined known devices operates rapidly, is simple and compact in structure and economical in manufacture and which further is free from the disadvantages set forth hereinabove.

Briefly stated, according to the invention, between the pressure source and the supply valve there is disposed a throttle member which affects the charging time and thus the injected fuel quantities and further, the supply valve is biased closed by means of a spring to a valve opening pressure which, taken together with the maximum charging pressure in the pump work chamber, is at most as large as the fuel supply pressure. The latter, in turn, is larger than the opening pressure of the fuel injection valve.

The fuel injection apparatus incorporating the aforeoutlined features may be of a very simple structure; it does not need additional servo pistons and it operates in a secure and rapid manner.

In addition, the fuel injection apparatus outlined above has the advantage that, since the charging periods are substantially longer (for example, seven-fold) than the periods of injection, a correspondingly greater accuracy in the metering of the injected fuel quantities is achieved. By virtue of the extended charging period even very small fuel quantities may be controlled with high accuracy which has not been possible heretofore.

Also, the disadvantage of unsatisfactory atomization of the fuel in case of small fuel quantities is eliminated with the fuel injection apparatus according to the invention. This applies in case of even smaller fuel injection quantities than those in known valves, since the pressure of the injected fuel quantities delivered by the pump piston to the injection valve is independent from the magnitude of the stroke. In this manner there is achieved even for the smallest injected fuel quantity a definite jet length with good atomization. The jet length for any injected fuel quantity is independent from the rpm.

Further, by virtue of the structure according to the invention the maximum permissible injected fuel quantity is determined by the maximum possible stroke of the pump piston. This feature ensures that the maximum permissible injected fuel quantity cannot be exceeded. This is particularly advantageous in diesel engines, in case the maximum permissible injected fuel quantity is identical to the full load fuel quantity. In this manner an emission, during excess fuel injection, of uncombusted pollutants prohibited with ever increasing severity by the clean air laws, is effectively prevented.

It is also an advantage of the structure according to the invention that by virtue of the throttle member, the charging period and thus the charging stroke at the maximum permissible rpm may be extended to the entire period between the terminal moment of the injection period of one work cycle and the starting moment of the injection period of the successive work cycle. In this manner an automatic safety control is achieved, because if the maximum rpm is exceeded, an automatic reduction of the injected fuel quantity is effected, since in case of such an rpm increase the charging period will no longer be sufficient for a complete charging of the pump work chamber.

The invention will be better understood as well as further objects and advantages become more apparent from the ensuing detailed specification of three exemplary embodiments taken in conjunction with the drawing.

BRIEF DESCRIPTION OF THE DRAWING

FIG. 1 is a longitudinal, partially schematic sectional view of the first embodiment of a pump-and-nozzle assembly of the fuel injection apparatus;

FIG. 2 is a longitudinal sectional view of the second embodiment of a pump-and-nozzle assembly;

FIG. 2a is a longitudinal sectional view of one part of FIG. 2 showing the pump piston in its lower dead center;

FIG. 3 is a longitudinal sectional view of the third embodiment of the pump-and-nozzle assembly and

FIG. 4 is a diagram illustrating the injected fuel quantities as a function of the injection, charging and control periods.

DESCRIPTION OF THE FIRST EMBODIMENT

Turning now to FIG. 1, the first embodiment of a pump-and-nozzle assembly 9 of a fuel injection apparatus is designed to serve an Otto engine wherein fuel is injected into the air intake tube. The pump-and-nozzle assembly 9 is tightened together by means of a nut sleeve 12 to constitute a unitary structure. The assembly 9 is formed of two structural groups 10 and 11. The first structural group 10 has a housing 13 which contains a solenoid control valve 16 inserted in a housing bore 14 which is an enlarged continuation of a transversal fuel inlet bore 15. The housing 13 further accommodates a hydraulically operated pump 17 formed substantially of a cylindrical pump piston 19 guided in a cylinder bore 18, a supply valve 21 and a throttle member 22.

The second structural group 11 comprises an injection nozzle 13 containing an outwardly opening fuel injection valve 24. The latter includes a valve body 25, a valve needle 26, a closing spring 27 and a spring seat disc 28 which is connected with the valve needle 26 in a known manner and which forms a support for the closing spring 27. The other support for the closing spring 27 is constituted by the valve body 25.

The supply valve 21 has a movable valve member 31 and a valve spring 32 which presses the valve member 31 against a valve seat 33 and which is seated in a depression 34 of the fuel injection nozzle 23.

The housing 13 of the pump-and-nozzle assembly 9 has a supply bore 35 to which there is coupled a supply conduit 36. The latter carries fuel which is admitted to the pump 17 at a supply pressure p.sub.Z from a pressure source 37. The latter, together with its accessory components, is of conventional structure (for example, a gear pump driven by the engine) and is therefore only symbolically illustrated.

The output pressure of the pressure source 37 is maintained by a pressure regulating valve 38 at the desired supply pressure, for example, at p.sub.Z = 5 kg/cm.sup.2. In order to compensate for pressure fluctuations, the pressure regulating valve 38 may be combined with a pressure accumulator, the structure of which, similarly to that of the pressure regulating valve 38, is of conventional nature and is therefore not illustrated or described in detail.

The solenoid control valve 16 which will be described in detail later, controls the flow of the fuel from or to a servo pressure chamber 39 located above the pump piston 19. The length of this chamber determines the maximum stroke H.sub.max and thereby the largest permissible injection quantity Q.sub.max. The maximum stroke H.sub.max may also be determined by varying the position of an upper abutment 41 or by shortening or lengthening the cylinder bore 18.

In the position shown, the pump piston 19 is, with its upper radial face 40, in contact with the upper abutment 41, while its lower radial face 42, remote from the abutment 41, bounds a pump work chamber 43 adjoining the cylinder bore 18. The pump work chamber 43 is bounded at its bottom by the injection nozzle 23, the upper radial face of which constitutes a lower abutment 47 for the pump piston 19. From the pump work chamber 43 there extends a connecting port 44 to the fuel injection valve 24. From the supply valve 21 fuel may enter into the pump work chamber 43 through a supply port 45. The throttle member 22 is threadedly engaged in a charging bore 46 which extends from the transversal bore 15 to the supply valve 21. The flow passage section of the throttle member 22 affects the supply speed of the fuel flowing to the supply valve 21 and thus to the pump work chamber 43. Stated differently, said flow passage section determines the charging period t.sub.F of the pump-and-nozzle assembly 9. By virtue of the charging period t.sub.F which, due to the throttle member 22, is very substantially lengthened (for example, seven-fold) with respect to the injection period t.sub.E, there is ensured a correspondingly more accurate metering of the fuel quantity Q. The lengthened charging period t.sub.F permits a very accurate control even of extremely small fuel quantities (smaller than 3 mm.sup.3 per stroke).

The solenoid control valve 16, shown in a simplified manner, is a conventional, pressure-equalized 3/2-way valve (disclosed for example, in German Published application DOS Pat. No. 1,934,212) operated by an electromagnet 51. The solenoid control valve 16 comprises a valve housing 52 and a sphere 53 which serves as the movable valve member. The sphere 53, in its position shown in FIG. 1 closes a valve seat 54 forming the opening of a port 55 which is in continuous communication with the transversal fuel inlet bore 15. Thus, in the position shown, the sphere 53 blocks the supply of the fuel from the pressure source 37 to the servo pressure chamber 39. Simultaneously, the servo pressure 39 is in communication with a return or discharge conduit 58 through a second valve seat 56 -- which in FIG. 1 is shown in its open position -- and through a discharge outlet 57. Through the conduit 58 which is coupled to the discharge outlet 57, the fuel returns in a depressurized condition to a fuel tank 59 from which the pressure source 37 draws fuel through a suction conduit 60.

The electromagnet 51 comprises an armature 61 guided in a bore 62 which is in axial alignment with the port 55. Urged by a spring 64, the armature, with an integral axial pin 63 presses the sphere 53 against the valve seat 54 when the electromagnet 51 is in a de-energized condition. To permit the use of a spring 64 of moderate strength, the solenoid control valve 16 is pressure-equalized by providing that the supply pressure p.sub.Z prevailing in the transversal bore 15 is, through a channel 65, communicated to a chamber 66 behind the armature 61. The surfaces of the sphere 53 and the armature 61 exposed to the pressure of the fuel are of identical magnitude, so that the forces derived from the fuel pressure and exerted on the sphere 53 in the opening and in the closing direction are also identical. Thus, the spring 64 has to supply a relatively small force merely to maintain the sphere 53 in contact with the valve seat 54.

As soon as the solenoid 67 of the electromagnet 51 is energized, for example, by means of an electronic control apparatus 68 (shown only symbolically), the force of the spring 64 is overcome by the electromagnetic forces and the armature 61 is moved to the right. The inflowing fuel passing through the now open valve seat 54 presses the sphere 53 against the second valve seat 56. Thus, the fuel may flow from the transversal bore 15 through the port 55 and the first valve seat 54 of the solenoid control valve 16 into the servo pressure chamber 39. The portions of the valve housing 52 which are exposed to different high pressures are separated from one another in the stepped bore 14 by sealing rings 69, 69a and 69b.

The moving components of the solenoid control valve 16 are of a small mass which permits almost delay-free switching operations. Such property is desirable for a rapid and accurate operation of the solenoid valve.

DESCRIPTION OF THE SECOND EMBODIMENT

The pump 17 and the injection nozzle 23 with the injection valve 24 as well as the supply valve 21 and the throttle member 22 do not have to be separately mounted as illustrated in FIG. 1, but may be, as illustrated in FIG. 2, which shows the second embodiment of a pump-and-nozzle assembly 9', combined into structural groups 71 and 72.

The structural group 71 comprises a pump housing 73 which is threadedly engaged in the housing 13' and which is provided with a cylinder bore 18' to slidably receive the pump piston 19'. The upward stroke of the latter is limited by an upper annular abutment 41'. The pump piston 19' is provided on its lateral face with an axial air bleeder groove 74 the length of which is so designed that in the lower dead center position (UT) of the pump piston 19 shown in FIG. 2a, in which the latter is in engagement with the lower abutment 47', it forms a throttle gap a with the upper boundary of the pump work chamber 43'. The throttle gap a is only a few tenths of a millimeter wide (for example, 0.1 - 0.2 mm). This small width however, is sufficient to permit an excape of the vapor or air bubbles from the pump work chamber 43' into the servo pressure chamber 39'. Therefrom, the bubbles are, in the closed position of the solenoid control valve 16, carried by the fuel which flows back to the fuel tank. In this manner, both the servo pressure chamber 39' and the pump work chamber 43' are de-aired. At the beginning of the charging (upward) stroke of the pump piston 19', the narrow throttle gap a and thus the air bleeded groove 74 are immediately closed, so that the latter practically does not affect the charging period t.sub.F. The lower radial face 42' of the pump piston 19' is adapted to seal the connecting bore 44' between the pump work chamber 43' and the injection valve 24 in the manner of a flat seat valve when the pump piston 19' is in its lowermost position. Thus, a passage of fuel to the injection valve 24 is prevented as long as the pump piston 19' is in engagement with the abutment 47'. The injection valve 24 is disposed in the lower part of the pump housing 73.

The second structural group 72 which is threadedly engaged in the supply bore 35' contains a supply valve 21' which corresponds to the supply valve 21 of FIG. 1 and which has a valve spring 32' and a throttle bore 75. The latter has the function of the throttle member 22 of the first embodiment.

The supply valve 21' communicates with the pump work chamber 43' through a supply bore 76 and a supply port 45'.

In the bore 14' which is an extension of the transversal bore 15', there is inserted, similarly to the first embodiment according to FIG. 1, a solenoid control valve 16 connected to a supply bore 35' and a return or discharge bore 57' which are coupled, respectively, to the supply conduit 36 and the return conduit 58. The pump-and-nozzle assembly 9' is associated with a pressure source and a fuel tank in a manner identical to the embodiment illustrated in FIG. 1.

DESCRIPTION OF THE THIRD EMBODIMENT

Turning now to the third embodiment illustrated in FIG. 3, there is shown a pump-and-nozzle assembly 9" adapted in particular for injecting fuel in diesel engines. This pump-and-nozzle assembly 9" differs from the pump-and-nozzle assembly 9' of FIG. 2 by a modified structural group 71" which has a longer pump housing 73" threadedly engaged in the housing 13". To the pump housing 73" there is tightened a fuel injection nozzle 23" by means of a sleeve nut 12" similarly to the embodiment illustrated in FIG. 1. The upper abutment 41" for the pump piston 19 is, similarly to FIG. 2, of annular configuration. The pump housing 73" has a substantially longer connecting bore 44" between its pump work chamber 43" and the fuel injection valve 24" than the housing 73 of the pump-and-nozzle assembly 9' of FIG. 2. The fuel injection nozzle 23" has a fuel injection valve 24", the valve body 25" of which contains a valve needle 26". The valve 24" is positioned in a bore 79 of the injection nozzle 23" by means of a support sleeve 78. The conduits 36 and 58 are coupled, respectively, to a pressure source and a fuel tank (neither shown) of known structure as shown in FIG. 1. The pressure source, however, has to be designed for a higher supply pressure p.sub.Z (for example, 200 kg/cm.sup.2) than in case of the first embodiment. Similarly to the pressure source, the supply valve 21" with the valve spring 32" and throttle bore 75" of the second structural group 72" as well as the fuel valve 24" have to be adapted to higher pressures. The solenoid control valve 16 is designed identically to that described in connection with FIG. 1 and is therefore not described here in detail. The pump-and-nozzle assembly 9" may also be used in case of correspondingly lower pressures for a direct injection into the combustion chambers of an Otto engine.

CONDITIONS FOR THE OPERATION AND GENERAL OPERATIONAL

FEATURES OF THE EMBODIMENTS

All three embodiments of the invention discussed hereinabove operate only if, according to the invention, the respective valve spring 32, 32' 32" of the supply valve 21, 21', 21" is biased to a valve opening pressure p.sub.V which, together with the maximum charging pressure p.sub.Fmax in the pump work chamber 43, 43', 43", is at most equal to the fuel supply pressure p.sub.Z. Furthermore, the charging pressure p.sub.Fmax has to remain smaller than the opening pressure p.sub.O of the injection valve 24, 24". For a fuel injection effected into the air intake tube of an Otto engine, the pump-and-nozzle assembly 9 and 9', according to FIGS. 1 or 2, respectively, operate in a reliable manner if, for example, p.sub.Z = 5 kg/cm.sup.2, p.sub.V = 2.4 kg/cm.sup.2, P.sub.Fmax = 2.6 kg/cm.sup.2, and p.sub.O = 3 kg/cm.sup.2. The pump-and-nozzle assembly 9" according to FIG. 3 operates well, for example, with pressures p.sub.Z = 200 kg/cm.sup.2, p.sub.V = 80 kg/cm.sup.2, p.sub.Fmax = 120 kg/cm.sup.2 and p.sub.O = 150 kg/cm.sup.2. The maximum charging pressure p.sub.Fmax in the pump work chamber 43, 43', 43", appears only if the pump piston 19, 19' dwells at least for a limited duration in contact with its upper abutment 41, 41', 41". In such a case the pressure p.sub.F in the pump work chamber 43, 43', 43" increases up to the moment when p.sub.F + p.sub. V reaches the value of p.sub.Z.

In FIG. 4 in the lower part of the diagram there is shown the course of the stroke H of the pump piston 19, 19', (and thus the injected fuel quantity Q) as a function of the charging, injection and control period (t.sub.F, t.sub.E, t.sub.S). The largest possible injection quantity Q.sub.max (highest point on curve A) is obtained, for example, at a stroke of H.sub.max and at a charging period of t.sub.F = 25 milliseconds (ms), while the associated injection period t.sub.E is 3.6 ms. The two periods correspond, respectively, to 315.degree. and 45.degree. of cam angle (NW), which together amount to 360.degree. cam angle, that is, one rotation of the cam shaft. In a four-cycle engine, one cam shaft revolution corresponds to two revolutions of the crank shaft, that is, 720.degree. of crank shaft angle (KW). The injection period and the charging period together (t.sub.E + t.sub.F) total, in the example according to the curve A, 28.6 ms which corresponds to the cycle period T of one work cycle of the engine in case of an engine rpm n - 4,200, since T = 2.60/4,200 = 2.360/6.4200 = 28.6 .times. 10.sup.-.sup.3 sec. = 28.6 ms. The equality T = t.sub.F + t.sub.E applies only if the end of the injection period and the start of the charging period occur at the same moment (t.sub.2).

The smaller injection quantity Q.sub.1 (for partial loads) is attained with a stroke H.sub.1 and with a course of injection as shown by the curve B drawn in broken lines. The associated charging period is t.sub.F1 and the corresponding injection period is t.sub.E1. In the illustration of the course of curve B according to FIG. 4 an engine rpm of n = 4,200 is assumed; a smaller rpm would yield a correspondingly larger cycle period T (not illustrated). Between the end of period t.sub.E1 and the beginning of period t.sub.F1 there extends the dwelling period t.sub.R1 during which the pump piston 19, 19' dwells in contact with its lower abutment. The cycle period T in this case is composed of t.sub.E1 + t.sub.R1 + t.sub.F1. The switching periods of the electromagnet 51 of the solenoid control valve 16 are shown by the solid-line curve C for the highest possible injection quantities Q.sub.max and by the broken-line curve D for partial load quantities Q.sub.1. At C.sub.1 and, respectively, D.sub.1, the valve 16 is in its closed position, whereas at C.sub.2 and, respectively, D.sub.2 it is in its open position. The beginning and the end of the energizing periods t.sub.S and t.sub.S1 determine the beginning moment t.sub.1 of the injection and the beginning moment t.sub.2 and t.sub.4, respectively of the charging stroke. The period between the two energizing periods t.sub.S and t.sub.S1 in which the solenoid 67 is in a de-energized condition and thus the valve is in its closed position C.sub.1 and D.sub.1, is identified as de-energized period t.sub.A, t.sub.A1. The moment of the termination of the injection is indicated respectively at t.sub.2 and t.sub.3 and is affected basically only by the readied injection quantity Q.sub.max and Q.sub.1, respectively, since the other influencing magnitudes, such as the supply pressure p.sub.Z and the characteristics of the injection valves 24 and 24" are constant. If desired, the fuel supply pressure p.sub.Z may be varied for altering the injection period within limits, for example, in an rpm-dependent manner.

OPERATION OF THE FIRST EMBODIMENT

In the description that follows, there will be set forth, with reference to FIGS. 1 and 4, the operation of the first embodiment for one work cycle T of the engine. All the other embodiments operate accordingly.

Prior to the beginning of the injection of the full load fuel quantity Q.sub.max (curves A and C), the pump piston 19 is disposed at H.sub.max at its upper abutment 41 (OT-position), since it just completed its charging stroke. At t.sub.1 the solenoid control valve 16 switches from the closed position C.sub.1 into the open position C.sub.2 whereby the sphere 53 quickly moves from the first valve seat 54 to the second valve seat 56 and the fuel, delivered at a supply pressure p.sub.Z from the pressure source 37, is admitted to the servo pressure chamber 39. There it exerts a force on the upper radial face 40 of the pump piston 19, driving the latter downwardly until the pump piston 19 engages, with its lower radial face 42, the lower abutment 47 at moment t.sub.2. During this downward motion, the pump piston 19 performs its maximum stroke H.sub.max and delivers the fuel prevailing in the pump work chamber 43 through the connecting bore 44 to the injection valve 24 of the injection nozzle 23. Since the fuel supply pressure p.sub.Z exerting a force on the upper radial face 40 of the pump piston 19 is greater than the opening pressure p.sub.O of the fuel valve 24, and further, since the supply valve 21 is closed, in the pump work chamber 43 there will now (that is, during the downward motion of the pump piston 19) prevail the supply pressure p.sub.Z. As a result, the injection nozzle 23 injects the fuel quantity Q.sub.max in a known manner into the air intake tube of the engine. At the end t.sub.2 of the injection period, the solenoid control valve 16, after an energized period t.sub.S, switches from its open position C.sub.2 back into its closed position C.sub.1. At that moment the sphere 53 shuts off the fuel supply to the servo pressure chamber 39 and depressurizes the latter by discharging the fuel through the now open second valve seat 56, the bore 62, the outlet bore 57 and the return conduit 58 into the fuel tank 59. At the same time, the pressure in the pump work chamber 43 drops very suddenly, the supply valve 21 opens and the fuel flows into the pump work chamber 43 with a charging pressure p.sub.F reduced by the throttle member 22 and the supply valve 21. This charging step occurs between moments t.sub.2 and t.sub.1 during the charging period t.sub.F, until, at t.sub.1, the solenoid control valve 16 again switches into its position C.sub.2 and the next cycle T is ready to start. In this case the charging period t.sub.F is identical to the deenergized condition t.sub.A of the solenoid control valve 16.

During delivery of a partial load fuel quantity Q.sub.1 according to the curves B and D, during the charging period t.sub.F1 between t.sub.4 and t.sub.1 there is readied, with a stroke H.sub.1, an injection quantity Q.sub.1 only. At moment t.sub.1, when the solenoid control valve 16 switches from D.sub.1 to D.sub.2, the injection stroke begins and terminates at moment t.sub.3. From this moment until the end t.sub.4 of the energized period t.sub.S1 of the solenoid 67, the pump piston 19 remains in contact with the lower abutment 47. This interval is the dwelling period t.sub.R1. As the solenoid valve 16 is switched from its open position D.sub.2 into its closed position D.sub.1 at moment t.sub.4, the charging stroke begins which takes place during the charging period t.sub.F1 until the moment t.sub.1, at t.sub.1 the successive injection starts and the operation described hereinbefore is repeated.

The different embodiments according to FIGS. 1 - 3 each show one pump-and-nozzle assembly 9, 9', 9" of the fuel injection apparatus. It is to be understood that in case of a multicylinder engine the fuel injection apparatus may have a greater number of pump-and-piston assemblies each of which may serve one or more cylinders.

Claims

What is claimed is:

1. A fuel injection pump-and-nozzle assembly forming part of a fuel injection apparatus serving an internal combustion engine, said assembly being of the type that has (a) a hydraulically operated pump piston having two opposite radial faces, (b) a pump work chamber bounded by one of said radial faces of said pump piston, (c) a servo pressure chamber bounded by the other of said radial faces of said pump piston, (d) a pressure source externally of said assembly for pressurizing fuel to a supply pressure, (e) first supply conduit means extending from said pressure source to said pump work chamber, (f) a supply valve disposed in said first supply conduit means, (g) second supply conduit means extending from said pressure source to said servo pressure chamber, (h) discharge or return conduit means extending from said servo pressure chamber, (i) a solenoid control valve for controlling the flow in said second supply conduit means and in said discharge conduit means for intermittently causing fuel to be admitted at said supply pressure from said pressure source into said servo pressure chamber to effect the delivery strokes of said pump piston, and (j) a fuel injection valve in communication with said pump work chamber, said fuel injection valve being biased closed with an opening pressure greater than the charging pressure prevailing in said pump work chamber at the end of the charging period, the improvement comprising

A. a throttle member disposed in said first supply conduit means between said pressure source and said supply valve, said throttle member being so dimensioned as to substantially lengthen the charging period, during which fuel is supplied from said pressure source to said pump work chamber, with respect to the injection period during which fuel is injected into the engine from said pump work chamber, and

B. spring means urging said supply valve towards a closed position against the flow of fuel from said pressure source, the supply valve closing pressure exerted by said spring means being so designed that taken together with the maximum charging pressure in the pump work chamber it at most equals said supply pressure, said supply pressure being greater than the opening pressure of said fuel injection valve, wherein the radial faces of said pump piston define identical pressure receiving areas, and wherein said pump piston is displaceable in order to effect the delivery strokes and the charging strokes of the pump-and-nozzle assembly by the pressurized fuel delivered into said servo pressure chamber and said pump work chamber.

2. An improvement as defined in claim 1, wherein the maximum permissible fuel injection quantity is determined by the maximum possible stroke of said pump piston.

3. An improvement as defined in claim 1, wherein said throttle member being so dimensioned that the charging period and the charging stroke at maximum permissible engine rpm is extensible to the entire period between the terminal moment of an injection period and the starting moment of the injection period of the successive work cycle.

4. An improvement as defined in claim 1, wherein said solenoid control valve is constituted by an electromagnetically operable 3/2-way valve including a sphere as its movable valve member.

5. An improvement as defined in claim 1, wherein said pump piston has an air bleeder groove establishing communication of narrow flow passage section between said pump work chamber and said servo pressure chamber when the pump piston is in its terminal position upon completing each delivery stroke.

6. A fuel injection pump-and-nozzle assembly forming part of a fuel injection apparatus serving an internal combustion engine, said assembly being of the type that has (a) a hydraulically operated pump piston having two opposite radial faces, (b) a pump work chamber bounded by one of said radial faces of said pump piston, (c) a servo pressure chamber bounded by the other of said radial faces of said pump piston, (d) a pressure source externally of said assembly for pressurizing fuel to a supply pressure, (e) first supply conduit means extending from said pressure source to said pump work chamber, (f) a supply valve disposed in said first supply conduit means, (g) second supply conduit means extending from said pressure source to said servo pressure chamber, (h) discharge or return conduit means extending from said servo pressure chamber, (i) a solenoid control valve for controlling the flow in said second supply conduit means and in said discharge conduit means for intermittently causing fuel to be admitted at said supply pressure from said pressure source into said servo pressure chamber to effect the delivery strokes of said pump piston, and (j) a fuel injection valve in communication with said pump work chamber, said fuel injection valve being biased closed with an opening pressure greater than the charging pressure prevailing in said pump work chamber at the end of the charging period, the improvement comprising

A. a throttle member disposed in said first supply conduit means between said pressure source and said supply valve, said throttle member being so dimensioned as to substantially lengthen the charging period, during which fuel is supplied from said pressure source to said pump work chamber, with respect to the injection period during which fuel is injected into the engine from said pump work chamber, and

B. spring means urging said supply valve towards a closed position against the flow of fuel from said pressure source, the supply valve closing pressure exerted by said spring means being so designed that taken together with the maximum charging pressure in the pump work chamber it at most equals said supply pressure, said supply pressure being greater than the opening pressure of said fuel injection valve, and wherein said pump piston has an air bleeder groove establishing communication of a narrow flow passage section between said pump work chamber and said servo pressure chamber when the pump piston is in its terminal position upon completing each delivery stroke.