the present invention will now be described in detail with the aid of several specific embodiments utilizing fig1 through 4 . a fuel injection system according to the present invention is represented in fig1 . the essential components are a conventional line injection pump 1 and an injection valve 2 . the line injection pump 1 is connected to the injection valve 2 via a first and a second injection line 3 and 4 whereby both injection lines are branching off a first distributor portion 5 which is connected to a pressure connection 6 of the line injection pump 1 . corresponding to a difference in the travelling time δt of a pressure wave originating at the line injection pump 1 , the first injection line 3 is selected to be shorter than the second injection line 4 . the injection lines 3 and 4 are connected down - stream to a second distributor portion 7 , whereby the first injection line 3 is equipped with a first check valve 8 and the second injection line 4 is equipped with a second check valve 9 . the check valves 8 and 9 are spring - loaded and prevent back - flow into the injection lines 3 and 4 . at the two branch locations of the two injection lines 3 and 4 in parallel , two by - pass valves 10 and 11 are provided which are electrically actuatable . the first by - pass valve 10 is connected directly to the first distributor portion 5 and the second by - pass valve 11 is connected to the second distributor portion 7 . a first and a second pressure - regulating valve 10a and 11a are connected in series to the by - pass valves 10 and 11 . they are spring - loaded and serve to maintain a preset pressure in the injection lines 3 and 4 . the injection valve 2 is also connected to the second distributor portion 7 . the by - pass valves 10 and 11 may be electrically actuated via respective first and second solenoids 12 and 13 . for the actuation of the solenoids 12 and 13 and the by - pass valves 10 and 11 interacting therewith , respective first and second switches 14 and 15 are provided . the injection valve 2 of the present invention represented in fig2 differs from conventional injection valves of the prior art . in conformity with the conventional main components of injection valves the injection valve 2 of fig2 comprises a valve holder 16 and a valve body 17 which are connected via a sleeve nut 18 . inside the valve body 17 a valve needle 19 is guided in an axially movable manner . the valve needle 19 is maintained in a closed position by a closing spring 20 . the valve needle 19 extends into a pressure chamber 21 and is formed as a differential piston having a pressure shoulder 22 . the pressure shoulder 22 serves as a transition to a greater diameter d 2 . the upper end of the valve needle 19 is formed as a piston 23 having a diameter d 3 . the valve needle seat 24 at the end opposite the piston 23 has a diameter d 1 . the diameter ratios are such that d 3 2 & lt ; d 2 2 - d 1 2 . the fuel is supplied via a connection 25 and a first bore 26 , which also represents the cylinder of the piston 23 . a second bore 27 branches off the first bore 26 and opens into the pressure chamber 21 . the mode of operation of the fuel injection system of the present invention will be described in the following paragraphs . when a piston of a respective pumping element of the line injection pump 1 begins to pump fuel , the two by - pass valves 10 , 11 are in an open position at first . due to the previous working cycle the injection lines 3 and 4 are still under pressure which corresponds to a pressure forced upon the injection lines 3 and 4 by the pressure regulating valves 10a and 11a during the opening phase oft he by - pass valves 10 , 11 during the final stage of the fuel pumping action . thus , the newly commencing fuel pumping action forces a discharge of fuel out of the pressure - regulating valve 10a , followed , with delay , by the opening of the second pressure - regulating valve 11a and discharge of fuel . the fuel which is expanded to atmospheric pressure by this step is recycled to the fuel tank . due to the electrical actuation of the solenoids 12 , 13 via the respective electronic switches 14 , 15 the closing of the by - pass valves 10 , 11 is achieved almost delay - free , whereby the pressure increase ( above the standing pressure ) in the injection lines 3 and 4 is initiated in order to prepare for the pre - injection . when the pressure build - up at the end of the first injection line 3 reaches a value that surmounts the opening pressure of the injection valve 2 , the fuel injection into the combustion chamber of the piston is initiated by the lifting process of the valve needle 19 ( fig2 ). since the speed of the pressure build - up and also the amount of the pressure gain in the area in front of the injection valve are basically determined by the additive effect of two overlapping pressure waves , one of which is travelling down - stream towards the valve holder and the other , due to reflection , is moving up - stream along the closed valve needle seat , a volume discharge is possible via the injection action into the combustion chamber , which yields the needed fuel for the pre - injection ( the end of which is determined by the opening of a by - pass valve due to the electrical actuation of a solenoid ), but also leads to an undesired pressure collapse . this pressure collapse , which preferably occurs at low engine revolutions , is caused by the slow displacement speed of the piston of the pumping element , but also by the effect of pressure increase as a result of two opposite pressure waves as described above . the pressure increasing effect results in a volume stream discharge during the pre - injection which , depending on the conditions of the pumping element such as , for example , the pre - stroke and the cam shape , may be greater than the fuel volume stream that is pumped from the displacement piston of the pumping element into the injection line at this very moment . when the fuel pressure in the valve holder , due to the termination of the overlap of the pressure waves ( the reflected pressure wave which was travelling up - stream in the injection line towards the line injection pump has passed and departed the end facing the valve holder ) and due to the relatively high volume discharge decreases below the level of the opening pressure of the injection valve the closing action of the injection valve is initiated . in practice , due to the injection line length , it takes too much time for the pressure to build up at the injection valve to reach a level at which the initiation ( by a control ) of the main injection , in a suitable time frame , may be carried out by closing the by - pass valve . this detrimental effect may be explained by the down - stream travelling time , at the speed of sound , of the pressure development ( in the upper portion of the first injection line ) which is undisturbed by the volume discharge . the present invention now provides a solution to the aforementioned problems . according to the present invention , a second pressure wave is generated with a delay corresponding to the desired time interval between the pre - injection and the main injection at the injection valve in order to yield the volume required for the main injection . this may be achieved by providing a second injection line 4 in parallel to the first injection line 3 . the difference in lengths is to be selected such that depending on the different travelling times of the pressure waves the desired delay between the pre - injection and the main injection is reached . in order to prevent the pressure wave , which arrives delayed at the end of injection line 4 , from introducing fuel into the end of the first injection line 3 , which is under reduced pressure and is facing the valve holder 16 , the injection line 3 is equipped with a first check valve 8 . in the same manner the injection of fuel into the end of the second injection line 4 by the first pressure wave that reaches the end of the injection line 3 first has to be prevented . thus , a second check valve 9 is provided at the end of injection line 4 facing the valve holder 16 . a further embodiment of the present invention is the combination of the by - pass valves 10 and 11 with respective pressure - regulating valves 10a and 11a . due to these pressure - regulating valves , the fuel pressure may only decrease to a standing pressure during the so - called outlet phase ( the pressure - regulating valves 10a and 11a are open ). the standing pressure is determined by pre - loaded springs and is set to an equal value at both pressure - regulating valves . the standing pressure also prevails when the pumping element is inactive . the introduction of a high standing pressure serves the purpose to make available portions of the fuel - supplying displacement movement of the pump element piston for generating an additional fuel volume to be injected by the injection valve . these portions were not unavailable before , because they only served to create the necessary compression volume during the build - up phase of the injection line pressure . such losses in the fuel supply may be tolerated during the pressure build - up phase for the preparation of the pre - injection , but during the subsequent , extremely fast pressure build - up in the time interval between the terminated pre - injection ( with subsequent pressure collapse ) and the starting main injection this creates an intolerable volume deficit of the injection volume needed . a high standing pressure therefore ascertains a fast pressure generation after the completion of the pre - injection , especially at low engine revolutions . it is expedient to select the standing pressure of the injection lines as high as possible due to the aforementioned reasons . a limiting factor , however , is the low opening pressure required for most of the conventional injection valves . increasing the opening pressure is met by the following problems as set forth in the form of a diagram represented in fig3 . the shown force - pressure diagram of a conventional injection valve represents the relation between the force at the valve needle shaft ( vertical axis ) and the pressure at the valve holder ( horizontal axis ). the straight line between the points o and b shows the course of the pressure - induced force at the valve needle body . at point b , a force f 2 is generated by the opening pressure p 0 which equals the opposing force of a spring which presses the valve needle into its seat . even an insignificant increase of the pressure in the injection lines lifts the valve needle , which causes the well - known increase of the pressure working surface of the valve needle . this increase in the pressure working surface ( at the same pressure ) leads to a sudden increase of the valve needle power ( straight line between b and c ) in the direction of the opening until the valve needle reaches its position of contact . when the pressure in the injection lines decreases to the closing pressure p s , the force of the valve needle follows the path of the straight line between c and d . at the point d the force of the valve needle falls below the spring force f 2 so that the valve needle falls back into its seat . this also causes a decrease of the hydraulic working surface at the valve needle ( straight line between d and e ). the impact energy which is generated in the form of elastic deforming work at the valve needle seat is determined by the spring force f 2 and the speed of the decrease of the pressure in the injection lines . when the decrease in pressure of the injection lines requires less time than needed by the valve needle to travel back from the abutment that limits the opening path into its seat ( which occurs occasionally ), then the elastic deformation work is exclusively determined by the spring force f 2 and the distance the valve needle has travelled . since modern injection systems already have a high impact speed of the valve needle there are concerns about further increasing the spring force f 2 which may result in exceeding the allowed surface pressure at the seat of the injection valve , a system was sought after that generates the required closing force but yields lesser impact energy when closing the valve . the injection valve of fig2 operates as follows : the fuel which is under high pressure enters the pressure chamber 21 via the pressure fast connection 25 that is attached to the valve holder 16 and via the second and third bores 27 and 27a . the fuel pressure acts on the pressure shoulder 22 in the form of a ring surface the size of which corresponds to the diameter difference between d 2 ( diameter of the shaft of the valve needle ) and d 1 ( diameter of the valve seat ). a further hydraulic working surface for the fuel pressure is the face of the piston 23 ( circular surface with a diameter d 3 ). the position as well as the movement of the valve needle 19 are controlled by a total of three forces acting directly upon the valve needle : the collaborating forces in the direction of the valve closing , i . e ., the force of the closing spring 20 and the force generated at the piston 23 by the pressure of the injection lines , and the third force acting in the opposite opening direction of the valve by acting upon the pressure shoulder 22 ( in the pressure chamber ) of the valve needle 19 . the results of these measures will be explained with the aid of the force - pressure diagram represented in fig4 . with the increase of the pressure in the injection lines the force , resulting from the product of the pressure in the injection line and the hydraulic working surface ( cross - section of the ring surface according to the diameter difference ( d 2 - d 1 )), acts upon the shaft of the valve needle 19 in the opening direction . this is represented in the diagram as a straight line between the points o and b . at the same time , the closing force , represented by the summation of the spring force f 1 of the closing spring 20 and the additional hydraulic force ( product from the pressure in the injection line and the circular cross - section d 3 ( piston 23 , fig2 )), increases to a value which at the point b is equal to the valve needle closing force . b therefore represents the intersection of the straight lines of the closing and opening forces . at point b , the resulting sign reversal of the summation force that is acting on the valve needle determines the opening pressure and the amount of the lifting force of the valve needle . the opening force is represented by the following equation : ## equ1 ## when the pressure in the injection line surmounts the opening pressure p 0 , an abrupt increase of the opening force due to the working surface increase of the valve needle to a value corresponding to the diameter d 2 takes place , resulting in an acceleration of the valve needle 19 ( opening ) until its shoulder contacts the intermediate piece ( straight line b - c ). the subsequent injection of the fuel into the combustion chamber causes the pressure in the injection lines to drop and after the injection is complete the opening force at the valve needle decreases according to a mathematical relation as represented by the straight line section between c and d . the point d represents the intersection between the course of the closing and the opening forces and determines the so - called closing pressure : ## equ2 ## the closing of the valve needle takes place as soon as the pressure drops below the closing pressure . when the valve needle falls back into its initial position , an immediate partial collapse of the opening force , according to the straight line section d - e , occurs which is due to the decrease in the hydraulic working surface . the working surface is reduced from ## equ3 ## the diagram shows clearly that , for identical opening pressures in a conventional and in the inventive system , in a system in which the closing pressure of the spring is supported by an additional hydraulic force ( generated with the assistance of the pressure of the injection line ), the force of the spring may be selected as low as desired when the hydraulic working surfaces are carefully selected and adjusted . this is in contrast to a conventional system in which the closing force is determined by the closing force of the spring alone ( fig3 ). accordingly , it is not necessarily required to reinforce the closing spring , when a higher opening pressure is required ( for example , for the introduction of a relatively high standing pressure at a different location in the system ) because a closing component that is controlled by the pressure in the injection lines may be introduced instead . it may also be noted that this method of increasing the actuating pressure also results in a favorable droplet size distribution during the fuel atomization which results in better fuel mixing properties for the purpose of reducing the amount of soot in the exhaust gases as well as improving the cold start properties . the present invention is , of course , in no way restricted to the specific disclosure of the specification , examples and drawings , but also encompasses any modifications within the scope of the appended claims .