Field of the Invention

The invention relates to fuel injection pumps for diesel engines and in particular to means for automatically advancing dynamic injection timing of such, pumps under cold conditions.

Rotary fuel injection pumps incorporate a transfer pump that is driven by the engine and which produces a fuel pressure (hereinafter referred to as a transfer pressure) within the pump housing that is proportional to engine speed. This transfer pressure is used to operate a hydraulic timing device that advances the dynamic injection timing with increasing engine speed. The timing device comprises a spring-loaded piston responsive to the transfer pressure and which is mechanically linked to the cam roller ring of a high pressure, rotary distributor-pump within the pump housing, thereby to vary the angular setting of the ring and thus the dynamic injection timing with engine speed.

Under cold conditions, the transfer pressure is boosted in order to operate the timing device and further advance dynamic injection timing with engine speed. This is achieved by selective operation of a pressure-relief valve that controls the setting of a pressure-regulating valve that regulates the transfer pressure. The pressure- regulating valve has a pressure control chamber that is connected via the pressure-relief valve to the fuel inlet supply. Under hot conditions, the pressure-relief valve is held open to connect the pressure control chamber directly to the fuel inlet pressure, but under cold operating conditions the pressure-relief valve is allowed to operate and generate a back pressure in the pressure control chamber that acts on the pressure-regulating valve to produce a

corresponding increase in the transfer pressure. Typically, the back pressure at engine idle speed is such as to produce a dynamic timing advance of about 5 degrees, giving an engine idle timing of 12 degrees advance under cold conditions compared with an engine idle timing of 7 degrees advance under hot running conditions.

Advancing the dynamic injection timing by this amount is sufficient to ensure that the engine starts and "runs-up" in an acceptable manner under cold conditions. However, as a consequence, the dynamic injection timing is increased throughout the speed range of the engine and can rise to a peak of 19 degrees advance in the mid-range 2000 to 2500 r.p. . As a result, there is a marked tendency for the engine to be noisy under cold operating conditions.

Disclosure of the Invention

An object of the present invention is to provide a fuel injection pump with cold timing control means that avoids the problem of excess noise under cold operating conditions.

According to the invention, a fuel injection pump for a diesel engine comprises a housing with a fuel inlet; a transfer pump mounted within the housing and adapted to be driven by the engine to pump fuel from the fuel inlet into the housing; a pressure-regulating valve that controls the transfer pressure of fuel within the housing produced by the transfer pump; a high pressure pump and distributor mounted within the housing and adapted to be driven by the engine to deliver fuel at high pressure from the housing to each of a plurality of fuel outlets in sequence; a timing device comprising a hydraulic actuator responsive to the transfer pressure and which is linked to the high pressure pump and distributor so as to advance the dynamic injection timing of the pump with increasing engine speed; and a cold condition actuator having a hot operating state for hot engine

conditions and a cold operating state for cold engine conditions and which controls the pressure regulating valve so as to increase the transfer pressure when in the cold operating state as compared with the transfer pressure when in the hot operating state, characterised in that the timing device includes restraining means controlled by the cold condition actuator so that when the actuator is in the cold operating state the restraining means becomes effective at a predetermined transfer pressure, corresponding to a require degree of advance of dynamic injection timing, and restrains further movement of the hydraulic actuator over a predetermined range of higher transfer pressure. Thus, the restraining means serves to stop movement of the hydraulic actuator during an intermediate phase, and thereby moderates the advance of dynamic injection timing to produce an advance/speed characteristic without excessive advance values.

Preferably, the restraining means comprises a resiliently- loaded stop which is engaged by the hydraulic actuator and which arrests movement of the actuator until the force exerted by the actuator on the stop overcomes the resilient loading of the latter.

Further, the stop is preferably hydraulically-loaded, and according to one embodiment comprises a counter-piston coaxially mounted with respect to the hydraulic actuator and resiliently loaded by hydraulic pressure of the fuel towards said actuator so as to assume a fixed position for engagement by the actuator. Increasing transfer pressure moves the actuator towards the counter-piston so that the two engage at said predetermined transfer pressure. Further movement of the actuator is then resisted until the transfer pressure increases sufficiently to overcome the hydraulic loading applied to the counter-piston by the fuel pressure. Preferably, the fuel pressure applied to the counter-piston

is a back-pressure produced in a pressure control chamber of the pressure-regulating valve by a pressure-relief valve. Thus, the actuator is held stationary until the back¬ pressure applied to the counter-piston is balanced out by a corresponding increase in the transfer pressure. Further increase of the transfer pressure then causes both the actuator and counter-piston to move again as one against the resilient-loading of the counter-piston.

The advantage of using the back-pressure to resiliently load the counter-piston is that it remains substantially constant with varying engine speed and is of a suitable magnitude greater than the fuel inlet pressure.

Description of the Drawings

The invention will now be described by way of example with reference to the accompanying drawings in which

Figure 1 is a schematic diagram of a fuel injection pump of a know type including a known hydraulic timing device.

Figure 2 is a schematic diagram of a fuel injection pump including a hydraulic timing device according to the invention.

Figure 3 is a graph illustrrating typical curves of dynamic injection timing against speed for the pumps of Figures 1 and 2,

Figure 4 is a graph illustrating typical curves of piston movement against speed for the pumps of Figures 1 and 2,

Figure 5 is an axial section through the hydraulic timing device of Figure 2,

Figure 6 is an end view of the counter-piston of the

hydraulic timing device of Figure S,

Figure 7 is a graph of dynamic injection timing against speed for various pumps as shown in Figure 2 with different dimensions of the components, and

Figure 8 is an axial section similar to Figure 5 showing an alternative form of hydraulic timing device according to the invention.

Best mode of carrrying out the Invention The know fuel injection pump illustrated in Figure 1 comprises a main pump housing 1 which incorporates a transfer pump 2 and a high pressure pump and distributor 3. The transfer pump 2 pumps fuel from the fuel inlets 4 into the housing 1, and a pressure-regulating valve 5 controls the tranfer pressure within the housing. The high pressure pump and distributor 3 delivers fuel from the housing to each of a plurality of fuel outlets 6 in sequence. The transfer pump 2 and the high pressure pump and distributor 3 are connected to a common input shaft 7 which is driven by the engine. The characteristic of the pump is such that the transfer pressure increases with engine speed.

A cam roller ring 8 and co-operating cam plate 9 form part of the high pressure pump and distributor 3, and a hydraulic timing device 10 serves to vary the angular setting of the ring 8 relative to the cam plate 9 so as to adjust the timing of the supply of fuel to the outlets 6, that is, to adjust the dynamic injection timing of the pump. The hydraulic timing device 10 comprises a spring-loaded piston 11 which is linked via a finger 12 to the cam roller ring 8 so that linear movement of the piston rotates the ring to adjust dynamic injection timing. A spring 13 engages one end of the piston 11 within a chamber connected to the fuel supply line 14 to inlet 4. The other end of the piston is

exposed to fuel within the housing 1 at transfer pressure so that an increase in transfer pressure with increasing engine speed serves to move the piston 11 against the spring 13 to adjust the ring 8 and advance the dynamic injection timing. Thus, with increasing engine speed the dynamic injection timing is advanced to compensate for the associated ignition delay.

The pressure -regulating valve 5 comprises a spring-loaded piston 15 within a valve housing 16 with a chamber 17 at one end connected to the outlet of the transfer pump 2. A port

18 opens through the side of the housing 16 into the chamber 17 and co-operates with the piston 15 to control venting of fuel to the fuel supply line 14, thereby regulating the transfer pressure within the housing 1. A restricted bore

19 through the piston 15 connects the chamber 17 to a pressure control chamber 20 at the opposite end of the piston containing the loading spring 21. Chamber 20 is connected via a pressure-relief valve 22 to the fuel supply line 14. The pressure-relief valve comprises a spring- loaded ball 23 that is loaded towards a valve seat 24 by a constant force spring 25.

An actuator 26, which may be electrically operated, is associated with the pressure-relief valve 22 so as to lift the ball 23 off the valve seat 24 for hot operating conditions. Under these conditions, the fuel flowing into the pressure control chamber 20 of the pressure-regulating valve 5 is free to recirculate through the pressure-relief valve 22 back to the fuel supply line 4. However, under cold operating conditions, the actuator 26 is de-energised to disengage the ball 23 and leave it free to operate under the control of spring 25. Thus, a back-pressure or balance pressure builds up in chamber 20 which increases the loading of the piston 15 of the pressure-regulating valve 5. Thus, the regulated pressure, that is, the transfer

pressure, is increased by a corresponding amount. This in turn produces an additional movement of the piston 11 of the timing device 10 and further advances the dynamic injection timing.

Figure 3 of the drawings shows typical curves of dynamic injection timing against engine speed for the known pump of Figure 1. The lower curve I is for hot operating conditions when the actuator 26 is operated to hold the pressure-relief valve 22 open. The upper curve II is for cold operating conditions when the actuator 26 is de-energised and the pressure-relief valve operates under spring control. It will be appreciated from these curves I and II that an appropriate level of advance of dynamic injection timing can be set for cold operating conditions with the engine idling, but that the levels of advance dynamic timing then produced as the engine speeds up are very high, rising to a maximum of about 19 degrees. At these advanced timings, the engine tends to be very noisy.

A fuel pump according to the invention is illustrated in Figure 2 and includes most of the components of the fuel injection pump of Figure 1, the same reference numbers being used in both drawings for common components. The main difference between the two pumps lies in the form of the timing control device 10. This includes the same piston 11, finger 12, and spring 13, but it also includes a hydraulically-operated stop 27 that is mounted on the pump housing 1 adjacent to the spring-loaded end of the piston 11.

With reference to Figures 5 and 6, the hydraulically- operated stop 27 comprises a housing 28 that is connected to the housing 1 and has an internal recess 29 supporting a counter-piston 30 in coaxial alignment with the piston 11. The recess 29 comprises two axially spaced cylindrical

chambers 32,33 separated by an intermediate transverse wall

31 having a pair of arcuate slots 34 formed through it around the circumference of the chambers. The counter- piston 30 comprises a cylindrical sleeve 35 with a transverse end wall 365 and a pair of arcuate lugs 37 formed as axially projecting extensions of the sleeve 35. The counter-piston 30 is located in the cylindrical chamber

32 as a slide fit with the lugs 37 projecting through the slots 34 towards the piston 11.

The spring 13 of the piston 11 has its outer end in engagement with the transverse wall 31. The outer end of the housing 28 is closed by a cover 38 and is provided with a fluid connector 39 for connection of chamber 32 to the conduit 40 connecting the pressure-regulating valve 5 to the pressure-relief valve 22. Another fluid connection 41 is provided to connect the chamber 33 to the fuel supply line 14.

When operating under hot conditions, with the actuator 26 holding the pressure-relief valve 22 open, fuel inlet pressure is applied to the chambers 32,33 both sides of the counter-piston 30. The latter is therefore substantially free to float axially within the housing 28 and does not restrain movement of the piston 11 which therefore operates as in the known pump described above.

When operating under cold conditions, with the actuator 26 de-energised and the pressure-relief valve operating under spring control, the back pressure or balance pressure generated in the conduit 40 is applied to the chamber 32 on the outer end of the counter-piston 30. The counter-piston is therefore urged axially into engagement witrh the transverse wall 31 with the lugs 37 projecting through the slots 34 into the chamber 33. However, the end of the lugs 37 are spaced axially away from the adjacent end of the

piston 11 and the latter is therefore free to move over the initial part of its travel as the transfer pressure increases from zero. The dynamic injection timing is therefore advanced with increasing engine speed as described above, until piston 11 engages the lugs 37. At this point, further movement of the piston 11 is restrained by the balance pressure applied to the counter-piston 30. If, however, engine speed continues to increase, the transfer pressure also increases until it overcomes the balance pressure on the counter-piston, whereupon the piston 11 and counter-piston 30 move as one under the effect of increasing transfer pressure against the action of spring 13.

A typical example of the movement of piston 11 with engine speed under hot and cold operating conditions, is illustrated in Figure 4. Curve I relates to hot conditions, and shows how the piston remains stationary with increasing speed until the transfer pressure balances the loading of spring 13 and moves the piston. Thereafter, the piston is moved uniformly with increasing speed over the whole of its range. Curve II relates to cold conditions, and shows how the piston moves initially with increasing speed until it contacts the lugs 37 of the counter-piston 30, after having travelled a distance of 2.5mm. The piston is then held stationary in this position until the transfer pressure builds up sufficiently to overcome the balance pressure. Thereafter, the piston 11 moves uniformly with increasing speed, as under hot conditions.

For purposes of comparison, a third curve III is shown on Figure 4 for the known pump of Figure 1 operating under cold conditions, and it will be seen that the piston 11 continues to move throughout the whole of its range with increasing speed and attains the maximum advance position at a much lower speed than the pump of Figure 2.

A similar comparison can also be made in Figure 3, in which curve III represents the curve of dynamic injection timing against speed corresponding to the curve II in Figure 4. It can be seen from curve III that the dynamic injection timing of the pump of Figure 2 operating under cold conditions does not rise to excessive levels like the known pump shown by curve II.

The operating characteristics of the pump of Figure 2 can be varied by varying the length of the lugs 37 and the diameter of the counter-piston 30. The length of the lugs 37 determine the position of the piston 11 when it engages the lugs 37 to be held stationary under cold operating conditions. The longer the lugs, the lower the speed at which the piston 11 is stopped. The diameter of the counter-piston 30 determines the effective force produced by the balance-pressure on the counter-piston that has to be overcome by the transfer pressure. Smaller diameters correspond to lower transfer pressures and thus lower speeds for movement of the piston from the stop position. These two affects are illustrated in Figure 7 for pistons with diameters D of 12, 16, 18, 20 and 24 mm., and stop positions S of the piston measured from the zero pressure, start position of 1.5, 2.0 and 2.5 mm. Curves for seven different combinations of piston diameter D and stop position S are shown in Figure 7. Further, a curve I shows the operating characteristic for hot conditions, and, for comparison purposes, a curve II shows the operating characteristic of the known pump of Figure 1 under cold conditions.

An alternative form of the hydaulically-operated stop 27 is shown in Figure 8 in which the intermediate transverse wall 31 of Figure 5 is replaced by an axially fixed control spigot 42 carrying a collar 43. The sleeve 30 of Figure 5 is replaced by a sleeve 44 having a base 45 with a central

hole 46 to receive the spigot 42 with the side wall of the sleeve surrounding the collar 43 and projecting towards the piston II.