JPS6157466B2 - - Google Patents

Description

[Detailed Description of the Invention] The present invention relates to an improvement of a distribution type fuel injection pump for a compression ignition engine, and more specifically, the present invention relates to an improvement of a distribution type fuel injection pump for a compression ignition engine, and more specifically, it injects a required amount of fuel in stages for the purpose of reducing exhaust harmful substances and combustion noise. Regarding equipment.

A distributed fuel injection pump as shown in FIG. 1 is known as a fuel injection pump that is well suited for relatively small high-speed compression ignition engines such as diesel engines for passenger cars.

To explain this, in the figure, fuel is sucked from the inlet 1 of the pump body by a feed pump 3 driven by a drive shaft 2 that rotates in synchronization with the engine, usually at half the speed of its crankshaft. .

The fuel discharged from the feed pump 3 is supplied to the pump chamber 5 inside the pump housing 5A after the supply pressure is controlled by the pressure regulating valve 4.

The fuel in the pump chamber 5 lubricates the operating parts and, at the same time, passes through the suction port 6 to the plunger pump 7 consisting of a plunger barrel 7A and a plunger 8.
sent to.

The plunger 8 has an eccentric disk 9
It is fixed to and is rotationally driven by the drive shaft 2 via a spline-shaped joint 2A.

The eccentric disk 9 is always pressed to the left in FIG. 1 by the spring 11. Also, this eccentric disk 9
has the same number of face cams 10 as engine cylinders (not shown), and rotates over the same number of rollers 13 as the face cams 10 disposed on a roller ring 12.

Therefore, the plunger 8 reciprocates in the axial direction within the plunger barrel 7A by a predetermined cam lift while rotating, and as a result of this rotational reciprocating movement, the air flows from the suction port 6 through the introduction groove 16 into the compression chamber 1.
The fuel sucked into the fuel pump 7 is forced into the distribution passage 14 through the passage 18, and then passes through the delivery valve 15 and is supplied to an injection nozzle (not shown).

Incidentally, the amount of fuel to be injected is determined by the position of the spill ring 20 that covers the spill port 19 formed in the plunger 8 in communication with the passage 18, that is, by changing the effective stroke of the plunger 8. 8 moves to the right, the spill port 19 escapes from the cover of the spill ring 20 and opens, releasing the high pressure fuel into the pump chamber 5 and ending the pumping.

The position of the spill ring 20 is fed back via an injection amount control device including a governor mechanism 21 and a start lever 31, which are driven in conjunction with the rotation of the drive shaft 2. That is, when the gear 33 attached to the governor shaft 32 rotates in response to the rotational force of the gear 2B fixed to the drive shaft 2, the flyweight holder 34, which rotates integrally with the gear 33, is rotated about a fulcrum. The flyweight 36 fixed to generates rotational centrifugal force and spreads around the fulcrum.

At this time, the flyweight 36 is attached to the governor sleeve 3 which is slidably fitted to the governor shaft 32.
A biasing force acts to press 7 and move it to the right.

Incidentally, the start lever 31 is attached to the collector lever 39 together with the tension lever 38 so as to be rotatable about a fulcrum A, and a spill ring 20 is attached to the extended tip of the start lever 31 via a ball joint 40. engaged.
The start lever 31 is connected to a control spring 41 provided to the tension lever 38.
By receiving a tensile force of through the start spring 42, the governor sleeve 37 is brought into close contact with the governor sleeve 37.

Therefore, when the governor sleeve 37 overcomes the tensile force of the control spring 41 and moves to the right,
The start lever 31 rotates around the fulcrum A, displacing the spill ring 20, and performing feedback control of the injection amount in accordance with the above-described procedure.

Note that the tensile force of the control spring 41 is adjusted by rotating a control lever 43 connected to an accelerator pedal (not shown), so that the injection amount can be controlled according to the load.

On the other hand, the timing of fuel injection is determined by the roller ring 12.
It is controlled by rotating the .

That is, since fuel is injected when the face cam 10 of the eccentric disk 9 rides on the roller 13, for example, if the roller ring 12 is rotated in the opposite direction to the rotation direction of the disk 9,
Since the time when the face cam 10 rides on the roller 13 becomes earlier, the fuel injection time becomes earlier.

In order to change the phase relative to the disk 9, a driving pin 23 is integrally fixed to the roller ring 12, and a timer piston 24 that can freely slide inside a cylinder 25, which will be described later, is connected to this driving pin 23.

Cylinder 2 on which the timer piston 24 slides
The fuel pressure of the pump chamber 5 is introduced to the right end high pressure chamber 26 of the pump chamber 5 through a passage 27, and the low pressure chamber 28 on the opposite side is communicated with the suction side of the feed pump 3 through a fuel relief hole 29. The timer piston 24 is displaced to a position where the pressure and the elastic force of the timer spring 37 are balanced. Note that FIG. 1 shows a state in which the axis of the piston 24 is rotated by 90 degrees, which actually coincides with the tangential direction of rotation of the roller ring 12.

Similarly, for convenience of explanation, the axis of the feed pump 3 is also shown rotated by 90 degrees.

Since the fuel pressure in the pump chamber 5 increases in proportion to the rotational speed of the feed pump 3, the piston 24 is pushed to the left as the engine rotational speed increases, thereby causing the piston 24 to move in the opposite direction to the rotation of the eccentric disk 9. It rotates the roller ring 12 and acts to relatively advance the injection timing.

In this way, the timer piston 24 that responds to the fuel pressure in the pump chamber constitutes a fuel injection timing control device that changes the rotational phase of the roller ring 12.

As described above, according to this distribution type fuel injection pump, it is possible to control the fuel injection amount and injection timing to be appropriate depending on the required operating state.

By the way, focusing on the action of the plunger pump 7, the stroke amount of the plunger 8 is always constant, and the effective stroke and injection amount are controlled by changing the timing at which the spill port 19 opens at the end of injection. The injection rate is approximately constant within the effective stroke range.

For this reason, the initial injection rate is relatively large, resulting in noticeable combustion noise and NOx in the early stages of combustion.
The problem arises that it is easy to cause

In other words, in compression ignition engines, there is generally an ignition delay phenomenon in which ignition occurs only after a short period of time after fuel is injected into the combustion chamber, and during this ignition delay the amount of fuel injected increases. This is because the initial pressure and temperature rise will be rapid.

On the other hand, for example, in Japanese Utility Model Publication No. 56-156459, two compression chambers of a plunger pump are provided coaxially, and the injection end timing of each is individually varied by two spill rings (ring sliders). A method has been disclosed in which noise is reduced by controlling the injection time regardless of the amount.
In this case, it is difficult to mechanically control the positions of the two spill rings in terms of accuracy, and in practice an expensive electrical control system would be required.

This invention was made by focusing on such conventional problems, and consists of defining two compression chambers coaxially in a plunger pump, and defining a pressure on the circumference of the suction port facing one of the compression chambers. This invention provides a multi-stage injection device for a distribution type fuel injection pump in which injection is started in stages by variably controlling the positions and shifting the compression start timing between each compression chamber according to the engine operating state. be.

In order to realize the above-mentioned staged injection, in the present invention, a fuel passage including a spill port and a distribution port is formed in the plunger corresponding to each of the two compression chambers, and the fuel passage from the distribution port is also formed in the plunger. A distribution passage is arranged around the plunger pump, supplying the corresponding engine cylinder via a check valve.

In addition, in order to shift the compression start timing phase between the two compression chambers, one of the compression chambers is surrounded by an annular auxiliary barrel that is rotatably supported around the plunger, and an annular auxiliary barrel is provided that faces the compression chamber. By rotating the barrel, the position on the circumference of the suction port formed in the barrel is controlled so as to open at the same time as the suction port facing the other compression chamber, thereby providing a difference in opening/closing timing between the suction port facing the other compression chamber and the suction port facing the other compression chamber.

The auxiliary barrel is provided with, for example, a hydraulic actuator that operates based on the pump chamber pressure, and is driven so that the suction port takes a predetermined position depending on the engine operating state.

The present invention will be explained below based on illustrated embodiments. Note that parts corresponding to those in FIG. 1 are designated by the same reference numerals.

In FIG. 2 or 3, reference numeral 8 denotes a plunger that rotates in synchronization with the engine rotation and reciprocates inside the plunger barrel 7A. Two compression chambers 17A and 17B are defined. However, the inner compression chamber 17A is located at the narrow diameter portion 8 of the plunger head.
Cylinder hole 51A into which A is slidably fitted
It is defined between the substantially annular auxiliary barrel 51 and the substantially annular auxiliary barrel 51.

The plunger 8 has inner and outer compression chambers 17A, 17.
Two mutually independent fuel passages 18A, 18B are formed corresponding to each of B. The two fuel passages 18A, 18B each include spill ports 19A, 19B that open facing the spill ring 20 and distribution ports 50A, 50B that open facing the inner peripheral surface of the plunger barrel 7A.

The plunger barrel 7A or pump body (not shown) has the two distribution ports 50A and 50.
A pair of distribution passages 14A, 14B open facing the discharge stroke area, corresponding to each of the distribution passages B
is formed. A set of distribution passages 14A, 14B
Check valves (or delivery valves) 15A and 15B are interposed in the middle of each, and they merge on the downstream side, and are connected to the injection nozzle (not shown) of a predetermined engine cylinder.
connected to. In addition, distribution passages 14A, 14B
For example, in the case of a six-cylinder engine, six sets are arranged so as to surround the plunger 8.

On the other hand, the auxiliary barrel 51 is the plunger barrel 7.
A substantially bottomed cylindrical cover plate 52 is fixed to the end face of the plunger 8 so as to be rotatable around the plunger 8, and as described above, a compression chamber is formed between it and the plunger narrow diameter portion 8A. 17A.

The auxiliary barrel 51 and the plunger barrel 7A are each formed with suction ports 6A and 6B that communicate with a fuel passage 6C from a pump chamber (not shown), and these ports 6A and 6B are formed at the shoulder of the plunger 8, respectively. Fuel is supplied to the compression chambers 17A and 17B through the inner and outer introduction grooves 16A and 16B. However, the connection portion between the fuel passage 6C and the port 6A is formed in an arc shape so as to maintain the connected state even if the auxiliary barrel 51 rotates to some extent (see FIG. 3). In addition, 53 and 54 are seal rings 55 and 56 for ensuring oil tightness between the auxiliary barrel 51, the plunger barrel 7A, and the cover plate 52.
is a plug screw for closing the open end of the auxiliary barrel cylinder hole 51A and a corresponding through hole 52A formed in the cover plate 52. By removing the two plug screws 55 and 56, the through hole 52A and the cylinder hole 5 are removed.
The lift of plunger 8 can be measured through 1A. Further, although only one inner and outer introduction groove 16A and 16B are shown in FIGS. 2 and 3, in a multi-cylinder engine, the number of guide grooves 16A and 16B are radially provided in accordance with the number of cylinders.

In this invention, the auxiliary barrel 51 is relatively rotated around the plunger 8 to control the compression start timing of the inner compression chamber 17A relative to the outer compression chamber 17B, thereby achieving a stepwise injection start. As an actuator for controlling the rotation of 51, a hydraulic cylinder device 60 is provided in this embodiment.

The hydraulic cylinder device 60 includes a cylinder portion 61 provided on the cover plate 52 in a tangential direction with respect to the auxiliary barrel 51, as shown in FIG.
The cylinder portion 61 includes a piston 62 that is slidably fitted in the cylinder portion 61, and a pin 63 that is implanted in the radial direction of the auxiliary barrel 51 is engaged with the piston 62.

The cylinder portion 61 has two pressure chambers 64 and 65 defined on both sides of the piston 62. The first pressure chamber 64 on the left side in the figure has the piston 62 in the second pressure chamber 64 on the right side.
A coil spring 66 that biases in the direction of 5 is interposed. The pressure on the suction side of the feed pump is introduced into the first pressure chamber 64, and the pump chamber pressure is introduced into the second pressure chamber 65, so that the position of the piston 62 is
It is determined according to the balance between the combined force of the pressure of No. 4 and the coil spring 66 and the pump chamber pressure of the second chamber 65.

Since the pump chamber pressure is proportional to the pump rotation, as the engine rotation increases, the piston 62 moves to the left and the auxiliary barrel 52 rotates clockwise. In order to perform control with higher precision or to enable position control that takes into account factors such as engine load, an electromagnetic relief valve 68 is interposed in the middle of the passage 67 that introduces the pump chamber pressure into the second pressure chamber 65. Ru.

The electromagnetic relief valve 68 controls the duty to adjust the pump chamber pressure to the first pressure chamber 65. For example, as the duty increases, the pressure in the first pressure chamber 65 is reduced and the pressure in the first pressure chamber 65 is reduced.
2 or the amount of clockwise rotation of the auxiliary barrel 51 can be reduced. In other words,
This makes it possible to control the position of the auxiliary barrel 51 in accordance with not only the engine rotation but also the load and engine characteristics.

Now, when the plunger 8 is rotating clockwise in FIG. 3, the pressure supplied to the first pressure chamber 65 of the hydraulic cylinder device 60 is increased to some extent and the auxiliary barrel 51 is rotated clockwise as described above. When the suction port 6B faces the outer compression chamber 17B,
The position of the suction port 6A of the inner compression chamber 17A is moved to the delayed side (as shown in the figure) compared to the above, and therefore, there is a difference in the timing of entering the compression stroke in the respective compression chambers 17A and 17B, and in this case, the inner compression Room 1
The injection start timing at 7A will be delayed.

In other words, when the outer introduction groove 16B closes the suction port 6A and starts pumping fuel, the inner introduction groove 1
6A still has the port 6A open, and only after the fuel in the outer compression chamber 17B has been injected to some extent through the passage 18B and the distribution passage 14B, the suction port 6A is closed and the pumping action in the inner compression chamber 17A is started. Ru.

Therefore, the fuel is injected in stages as shown in FIG. 4 by the delay in the start of pressure feeding in the inner compression chamber 17A.
Only a relatively small amount of fuel will be injected due to the pumping action at B. In other words, the initial injection rate becomes smaller.

As a result, the amount of fuel supplied to the engine combustion chamber during the ignition delay, that is, the amount of fuel combusted in a premixed state, is reduced, so the rise in pressure and temperature at the beginning of combustion is relatively slow, resulting in NOx This reduces the amount of gas generated and combustion noise.

Furthermore, the injection start timing in the inner compression chamber 17A can be retarded by variable control of the positions of the auxiliary barrel 51 and the suction port 6A by adjusting the pump chamber pressure to the hydraulic cylinder device 60. It is possible to control the optimum initial injection rate depending on the state, characteristics, application, etc.

In addition, since check valves 15A and 15B are interposed in the middle of the two distribution passages 14A and 14B, respectively, in this case, the fuel in the outer compression chamber 17B, which has entered the injection process first, passes through the passage 14A and flows into the inner side. There is no possibility that the fuel will flow back into the compression chamber 17A (if the injection start timing of one of the two compression chambers 17A and 17B, for example, the inner compression chamber 17A, is set to always be later than that of the other compression chamber 17B) teeth,
Check valve 1 of distribution passage 14B to which injection pressure acts first
5B may be omitted. ) Also, the end timing of injection in the inner and outer compression chambers 17A and 17B is determined by the respective spill ports 19A and 19B.
Since the timing is determined by the time when the spill leaves the spill ring 20, if the two spill ports 19A and 19B are opened at the same position in the axial direction of the plunger, the injection ends at the same time.

FIG. 5 shows a second embodiment of the present invention, in which an auxiliary barrel 51 is provided to surround an outer compression chamber 17B, and the injection start timing in the outer compression chamber 17B is variably controlled. . Since the other points are the same as those in FIG. 2, substantially the same parts are given the same reference numerals and their explanation will be omitted.

As explained above, according to the present invention, one of the two compression chambers coaxially defined in the plunger pump is surrounded by an auxiliary barrel rotatably supported around the plunger, and the suction port is By variably controlling the position on the circumference and shifting the injection start timing in the compression chamber relative to the other compression chamber, the fuel injection amount is increased in stages to achieve the optimal initial injection rate according to the engine operating state. Since we made it possible to control
Engine combustion noise and NOx can be effectively reduced.

[Brief explanation of the drawing]

FIG. 1 is a longitudinal sectional view of a conventional example. FIG. 2 is a longitudinal cross-sectional view of a main part of the first embodiment of the present invention, and FIG. 3 is a cross-sectional view taken along the - line. FIG. 4 is an explanatory diagram of the operation of the first embodiment. FIG. 5 is a vertical cross-sectional view of a main part of a second embodiment of the invention. 6A, 6B... Suction port, 7A... Plunger barrel, 8... Plunger, 14A, 14B...
...Distribution passage, 15A, 15B...Check valve, 16
A, 16B...Introduction groove, 17A, 17B...Compression chamber, 18A, 18B...Fuel passage, 19A, 19
B... Spill port, 20... Spill ring, 5
0A, 50B...Distribution port, 51...Auxiliary barrel, 60...Hydraulic cylinder device (actuator).

Claims (1)

[Claims] 1. A plunger pump that rotates and reciprocates in synchronization with the engine, a spill ring that is slidably fitted to the plunger and controls the opening timing of the spill port leading to the plunger compression chamber, and A distribution type fuel injection pump comprising an injection amount control device that controls the position of a spill ring, and an injection timing control device that also controls the phase of a roller ring that imparts reciprocating motion to the plunger. Two compression chambers are defined, and a fuel passage provided with a spill port and a distribution port is connected to the plunger corresponding to each of the two compression chambers, and fuel from the distribution port is connected to the corresponding engine cylinder. Supply distribution passages are each formed around the plunger pump, and one of the two compression chambers is rotatably supported around the plunger and includes a suction port opening facing the compression chamber. A distribution type fuel injection pump, characterized in that it is surrounded by a substantially annular auxiliary barrel and is provided with an actuator that drives the auxiliary barrel and controls the circumferential position of the suction port according to engine operating conditions. Multi-stage injection device.