JP2002031018A - High pressure fuel pump - Google Patents

Abstract

(57) [Summary] A single-cylinder plunger-type high-pressure fuel pump has a reduced volumetric efficiency especially at high rotation speed due to a delay in closing a suction / discharge valve. A cam has a cam profile in which a maximum peak position of a cam speed is approximately 1/3 to 1/2 of a discharge stroke section in a discharge stroke of a plunger from a bottom dead center to a top dead center. . As a result, the theoretical flow rate in the section where the suction valve is closed and the discharge becomes invalid is reduced. In addition, since the pressure rise rate of the fuel in the pressurizing chamber 12 is sufficiently secured,
The discharge valve opens promptly, and the section in which invalid discharge is performed can be sufficiently shortened. Therefore, as a result of an increase in the theoretical flow rate in a section where the effective discharge is performed after the discharge valve is opened, the discharge flow rate is increased, and the volumetric efficiency particularly at high rotation can be improved. By appropriately suppressing the rise of the cam speed at the start of the discharge stroke, the theoretical flow rate in a section where invalid discharge is performed due to a delay in closing the suction valve is reduced, and the theoretical flow rate in a section where effective discharge is performed thereafter is increased. For this reason, the discharge flow rate increases, and the volumetric efficiency especially at high rotation can be improved.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a high-pressure fuel pump for supplying fuel to a fuel injection valve of an automobile engine at a high pressure.

[0002]

2. Description of the Related Art A conventional high-pressure fuel pump is described, for example, in Japanese Patent Application Laid-Open No. 2-176158. The high-pressure fuel pump operates along a cam profile of a cam to suck fuel into the pressurized chamber and discharge the engine to a plunger, and a solenoid valve operated by a solenoid to supply fuel to the pressurized chamber. And This solenoid valve is operated by a solenoid so as to be able to cope with a case where an automobile engine rotates at a high speed.

In this conventional high-pressure fuel pump, the maximum pressure point of the suction and discharge strokes (discharge stroke) comes in the first half while the cam profile makes one rotation. This is to increase the fuel pressure in the pressurization chamber in a short time and pressurize the solenoid valve to a pressure that can keep the solenoid valve closed, thereby responding to high-speed running of the vehicle and shortening the time until the solenoid valve closes. In this case, a large amount of high-pressure fuel can be discharged.

[0004]

This type of fuel pump has two types of means for opening and closing a suction valve for sucking and discharging fuel: a type operated by a solenoid as in the above-mentioned prior art, and a valve operated by a spring. There is a type that presses against a seat and opens and closes a valve with a pressure difference in a pressurized chamber.

[0005] Of these types, in the case of a valve using a spring, in particular, when the engine of an automobile equipped with a fuel pump rotates at a high speed, the rotational speed of the cam and the opening / closing speed of the intake valve do not match, and the intake valve side There is a problem in that the amount of fuel flowing backward from the tank increases. In this case, it is conceivable to increase the size of the pump in order to secure a discharge amount equal to or higher than the reverse flow rate of the fuel.However, when the size of the pump increases, the load on the engine that drives the pump increases, and fuel efficiency decreases. There's a problem.

The cam profile of the conventional high-pressure fuel pump is effective when combined with a solenoid-operated valve, that is, a so-called solenoid valve. However, the cam profile is combined with a valve using a spring. It is unlikely that they will have the same effect.

That is, the fuel pump of the prior art using an electromagnetic valve can close the valve at an arbitrary timing by inputting a signal to the electromagnetic valve, so that the valve closing delay can be eliminated. When the valves are combined, when the cam speed is at the maximum, the operation of the valve cannot follow the cam speed, and the reverse flow rate of the fuel increases.

[0008] It is an object of the present invention to provide a high-pressure fuel pump having a reduced backflow of fuel.

[0009]

The object of the present invention is to provide a pressurized chamber communicating with a fuel intake passage and a discharge passage, a plunger reciprocating in the pressurized chamber by the shape of a rotating cam,
In a high-pressure fuel pump including a suction valve that opens and closes the suction passage and a discharge valve that opens and closes the discharge passage, the suction valve is opened immediately after the plunger starts a discharge stroke from bottom dead center to top dead center. This is achieved by suppressing the rising speed of the plunger while the cam is shaped to increase the speed after the suction valve is closed.

A pressurizing chamber communicating with a fuel intake passage and a discharge passage; a plunger reciprocating in the pressurized chamber by the shape of a rotating cam; a suction valve for opening and closing the suction passage; A high-pressure fuel pump having a discharge valve that opens and closes a passage, wherein the high-pressure fuel pump includes an elastic member that presses the plunger in opposition to the urging force of the cam, and the plunger is disposed in a discharge stroke from a bottom dead center to a top dead center. Following the first cam portion, in which the cam acceleration from the bottom dead center is suppressed to a small positive value, a second cam portion is provided, in which the cam acceleration is largely increased and then gradually reduced to a negative acceleration. A third cam having a negative cam acceleration following the second cam portion up to the top dead center at substantially the same value as a limit acceleration at which the cam and the plunger do not separate from each other even at the maximum reciprocating frequency of the plunger. In the provided, and the intake stroke of the plunger reaches the bottom dead center from the top dead center is achieved by having a cam profile with a target acceleration curve with respect to the cam acceleration and top dead center of the discharge stroke.

Further, the cam acceleration is achieved by providing a cam profile that makes the cam acceleration at the bottom dead center and the top dead center substantially zero.

A pressurizing chamber communicating with a fuel intake passage and a discharge passage; a plunger reciprocating in the pressurized chamber by the shape of a rotating cam; a suction valve for opening and closing the suction passage; A high-pressure fuel pump having a discharge valve for opening and closing a passage, comprising: an elastic member that presses the plunger in opposition to the urging force of the cam, wherein the cam discharges the plunger from a bottom dead center to a top dead center. Following the first cam portion, in which the cam acceleration from the bottom dead center in the stroke is a substantially constant positive value, a second cam portion is provided, in which the cam acceleration is gradually reduced to monotonically decrease to a negative acceleration. Second
The negative cam acceleration up to the top dead center following the cam portion is set to a substantially constant negative value, and the negative acceleration is substantially the same as the limit acceleration at which the cam and the plunger do not separate at the maximum reciprocating frequency of the plunger. A cam profile having a target acceleration curve with respect to the cam stroke of the discharge stroke and a top dead center in a suction stroke in which the plunger is moved from a top dead center to a bottom dead center. Is achieved.

Further, the cam is achieved by setting the maximum peak position of the cam speed to be approximately 1/3 to 1/2 of the discharge stroke section.

Further, the pump is achieved by not providing an urging means such as an electromagnetic actuator capable of opening or closing the suction valve at an arbitrary timing.

[0015]

DESCRIPTION OF THE PREFERRED EMBODIMENTS In the present invention, a high-pressure fuel pump uses a valve whose intake valve opens and closes in synchronization with the reciprocating movement of a plunger, a so-called check valve. With this valve, the valve cannot be closed at an arbitrary timing, so that the operation of the plunger involves a slight valve closing delay. For example, even if the plunger enters the discharge stroke, the suction valve does not close immediately,
During this time, the fuel pressurized by the plunger flows back out of the pressurization chamber through the suction valve. The amount of discharge of the pump is reduced by the amount of the backflow, and the volumetric efficiency particularly at high rotation tends to decrease.

Therefore, if the above-mentioned prior art cam is simply used in the high-pressure fuel pump of the present invention, the cam speed in the invalid section until the suction valve closes is increased, so that the reverse flow to the suction side is increased, and the discharge amount is rather increased. May be reduced.

First, a general high-pressure fuel pump will be described with reference to the drawings.

FIG. 1A is a longitudinal sectional view showing a high-pressure fuel pump. FIG. 1B is a partial cross-sectional view of the engine showing a state in which the high-pressure fuel pump is attached to the engine. FIG. 2 is a configuration diagram showing a fuel supply device for an engine using a high-pressure fuel pump.

In FIG. 1A, a pump body 1 includes:
A fuel intake passage 10, a discharge passage 11, and a pressurizing chamber 12 are formed. The plunger 2, which is a pressure member, is slidably held in the pressure chamber 12 by a cam 100. 3
Is a lifter for holding the spring 4 and the tip of the plunger 2. This lifter 3 is always in contact with the cam profile of the cam 100. 20 is plunger 2
And a seal member for sealing the inside of the pressurizing chamber 12 of the pump body 1. In the suction passage 10 and the discharge passage 11,
A suction valve 5 and a discharge valve 6 are provided.
The check valve is pressed against the valve seat at a and 6a to restrict the flow direction of the fuel. The pump body 1 composed of these components is generically called a high-pressure fuel pump 101.

In FIG. 1B, reference numeral 73 denotes an engine cover which houses a piston, an engine cam and the like. Reference numeral 72 denotes an engine camshaft.
0 is directly connected. The high-pressure fuel pump 101 is attached to a part of the engine cover 73.

As described above, when the engine camshaft 72 rotates, the cam 100 rotates, and the high-pressure fuel pump 1
The plunger 2 in 01 moves up and down. Such a high-pressure fuel pump is called a single cylinder plunger pump.

In FIG. 2, the fuel indicated by the arrow is regulated to a constant pressure by a low-pressure pressure regulator 52.
It is guided from the tank 50 to the fuel inlet of the high-pressure fuel pump 101 via the low-pressure pump 51 and the low-pressure pipe 9.
The fuel sucked into the high-pressure fuel pump 101 is pressurized by the rotation of a cam 100 provided in the engine 71, and is fed to the common rail 53 from a fuel outlet. The fuel pressure in the common rail 53 is maintained at a substantially constant pressure by a high pressure regulator 55. The injectors 54 are mounted on the common rail 53 according to the number of cylinders of the engine. The injector 54 injects fuel into the engine 71 according to a signal from an engine control unit (ECU). Reference numeral 56 denotes a pressure sensor, which detects the pressure of the common rail 53 and controls the opening and closing of the suction valve by the ECU according to the pressure.

The operation of the high-pressure fuel pump having the above configuration will be described.

In FIG. 1A, the lifter 3 provided at the lower end of the plunger 2 is pressed against the cam 100 by the spring 4, so that the plunger 2 is rotated by an engine camshaft or the like. The reciprocating motion follows the cam profile (outer shape) and changes the volume in the pressurizing chamber 12.

Each of the suction valve 5 and the discharge valve 6 is an automatic valve that opens and closes in synchronization with the reciprocating motion of the plunger 2.
During the suction stroke, the discharge valve 6 closes due to a decrease in the fuel pressure in the pressurizing chamber 12, and the suction valve 5 opens to suck the fuel into the pressurizing chamber 12. During the discharge stroke, the suction valve 5 closes and the discharge valve 6 opens due to an increase in the fuel pressure in the pressurizing chamber 12, and FIG.
Is fed to the common rail 53 shown in FIG.

The single cylinder plunger pump described above is
It has the advantages of simple structure and low cost.
Since the discharge valve 6 is held only by the pressing force of the spring 4,
Both valve opening and valve closing are delayed, and both cause a decrease in the discharge amount.

That is, a delay is caused by the moving distance due to the opening and closing of the valve. For example, even when the plunger 2 enters the discharge stroke from the suction stroke, the suction valve 5 is not closed immediately, and the fuel flows back from the pressurizing chamber 12 to the suction passage 10 side, so that the discharge flow rate decreases. In particular, when the vehicle is running at high speed, the cam rotates at a high speed, so that the vertical movement cycle of the plunger is shortened, the effect of the delay in closing the valve becomes more remarkable, and the volume efficiency is reduced.

If the volumetric efficiency is low, the displacement of the pump itself must be increased in order to secure the maximum discharge rate, and therefore the driving torque increases and the fuel efficiency of the engine deteriorates. Therefore, improvement of the volumetric efficiency at the time of high rotation is one of the important issues for the single cylinder plunger pump.

Therefore, in the present invention, while the suction valve 5 is still open immediately after the start of the discharge stroke, the rising speed of the plunger 2 is suppressed to reduce the flow rate flowing backward from the suction valve 5, and
After the is closed, the ascending speed of the plunger 2 is increased.

Therefore, according to the present invention, as shown in FIG. 3, the cam profile of the cam 100 is such that the maximum peak of the cam speed is substantially between 1/3 and 1/2 of the discharge stroke section. is there.

FIG. 3 is a diagram showing an example of a three-mounting cam which reciprocates the plunger three times with one rotation of the cam provided with the present invention.

In FIG. 3, assuming that the bottom dead center of the plunger 2 is a cam rotation angle of 0 °, in this case, about 20 ° to 30 °,
In the case of a double cam, the maximum peak of the cam speed is set at a position of about 30 ° to 45 °. As a result, the rise of the cam speed is moderately suppressed, so that the theoretical flow rate (= plunger cross-sectional area ×
Speed) decreases, and the theoretical flow rate in the section where effective discharge is performed after the discharge valve is opened increases. In addition, since the fuel pressure increasing speed is sufficiently ensured, the discharge valve is quickly opened, and the section in which the invalid discharge is performed can be sufficiently shortened. Therefore, the discharge flow rate is increased, and the volumetric efficiency particularly at high rotation can be improved.

Hereinafter, the present invention will be described in more detail with reference to FIGS. 4, 5 and 6.

FIG. 4 shows a comparison between the cam profile shown in FIG. 3 and other cam profiles.

In FIG. 4, a thick line indicates the cam A of the present invention, and a thin line indicates that the speed peak is 1 / of the discharge stroke section.
The cam B before 3 is shown, and the dotted line shows a cam C whose sine wave cam has a speed peak just half of the discharge stroke section.

Since the absolute value of the negative cam acceleration near the top dead center of the sine wave of the cam C is large, the inertial force of the reciprocating portion including the plunger 2 also increases. The inertial force becomes larger than the contact force between the lifters 3, and jumping occurs in which the two are separated from each other.

Jumping is to be avoided because it causes noise and lowers durability. In fact, the cam C cannot be used. In order to avoid jumping by using the cam C, the spring force must be increased, which causes adverse effects such as a decrease in durability due to an increase in Hertz stress on the contact surface and a decrease in fuel efficiency due to an increase in drive torque.

Next, the discharge flow rate characteristics of the pump when the cams A and B are used are shown in FIG.

In FIG. 5, the horizontal axis represents the pump rotation speed, and the vertical axis represents the time-averaged discharge flow rate. As described above, the single cylinder plunger pump shows a tendency that the volume efficiency decreases as the rotation speed increases, and the actual discharge flow rate decreases with respect to the theoretical flow rate.
The cam A according to the present invention has a smaller flow rate decrease than the cam B, and can maintain high volumetric efficiency up to high rotation. The reason will be described with reference to FIG.

FIG. 6 shows a time history waveform of each part of the pump at the time of high rotation, in which the case of the cam A is indicated by a thick line, and the case of the cam B is indicated by a thin line. As shown in the drawing, the suction valve 5 is still open even when the cam enters the discharge stroke from the suction stroke. Therefore, the flow rate corresponding to the pressure applied by the plunger 2 during this time passes through the suction valve 5 to the suction passage 10 side. Backflow.
For this reason, the discharge flow rate decreases.

Further, there is a delay from when the suction valve 5 is closed to when the discharge valve 6 is opened, and the total of both delays is an invalid section which does not contribute to the discharge. At the time of high rotation, the time per one stroke becomes short, so this invalid section relatively increases, and the volumetric efficiency decreases. The reason why the closing of the suction valve 5 is delayed is that even when the plunger 2 starts to descend, the pressure chamber 12
This is because the pressure of the suction valve does not drop and it takes time to move the suction valve itself from the open state to the closed state. The reason why the opening of the discharge valve 6 is delayed is that it takes time for the pressure of the pressurizing chamber to rise to the pressure of the discharge passage 11 downstream of the discharge valve.

Here, since the cam A of the present invention has a speed peak behind the cam B, the speed in the invalid section is kept low. Therefore, the theoretical flow rate in the invalid section proportional to this speed, that is, the invalid flow rate flowing back to the suction side decreases, and conversely, the theoretical flow rate increases in the section where the effective discharge is performed after the discharge valve is opened. The discharge flow rate of the cam A increases by the difference between the invalid flow rates. The difference between the two is small when the rotation speed is short and the invalid section is short, but when the rotation speed is high and the number of invalid sections increases, the difference gradually increases, resulting in a flow characteristic difference as shown in FIG.

From the above, it can be seen that, in order to improve the volumetric efficiency, it is effective to arrange the peak position of the cam speed as rearward as possible and suppress the cam speed in the invalid section.

However, if the peak position is set too far behind, the rise in the pressure of the pressurizing chamber after closing of the suction valve becomes slow, so that the opening delay of the discharge valve becomes particularly large. As a result, the invalid section itself becomes longer, and conversely, the discharge flow rate is reduced.

That is, the peak position of the cam speed may not be too early or too late, and may be approximately 1/3 to 1 of the discharge stroke section.
The range of / 2 is most suitable. Speed peak is 1/3
Before the position, the rise of the cam speed becomes too fast like the cam B, so that the invalid flow rate increases and the discharge flow rate decreases. Conversely, if it is less than 1/2, the rise in fuel pressure in the pressurized chamber is delayed, the delay in closing the suction valve and the delay in opening the discharge valve increase, and the invalid section itself becomes longer, which also leads to a decrease in the discharge flow rate. . 1/3 to 1/2 is an appropriate range for achieving both reduction of the invalid flow rate and shortening of the invalid section.

Since the cam speed is obtained by integrating the cam acceleration, it is necessary to appropriately set the cam acceleration in order to operate the cam speed. As mentioned above,
Although the sine wave cam in FIG. 4 was inappropriate in that jumping occurred, it was difficult to increase the absolute value of the negative cam acceleration near the top dead center as in the case of the sine wave cam from the following points. It is.

First, near the top dead center, the force of the return spring 4 of the plunger becomes the strongest, so that the Hertz stress acting on the cam surface also becomes large. To lower the Hertz stress,
Either lower the spring force or increase the radius of curvature of the cam near the top dead center.First, to reduce the spring force, reduce the negative absolute value of the cam acceleration to reduce the inertia force of the reciprocating part. Must be reduced so that the cam 100 and the lifter 3 do not separate from each other even with a weak spring force. Also, in order to increase the radius of curvature, it is necessary to reduce the negative absolute value of the cam acceleration or increase the base circle of the cam. Of these, the base circle cannot be so large from the standpoint of miniaturization, so in order to keep the Hertz stress within a certain allowable value and ensure durability, the negative absolute value of the cam acceleration near the top dead center should be small. I have to do it.

For this reason, it is difficult to reduce the cam speed in a short section, and the deceleration section becomes longer (the acceleration section becomes shorter). As a result, the maximum peak position of the cam speed becomes earlier than １ ／ of the discharge stroke section. It tends to be.

For this reason, in the cam shown in FIG. 3, the cam acceleration from the bottom dead center is suppressed to a small positive value in order to set the maximum peak position of the cam speed to 1/3 to 1/2 of the discharge stroke section. Following the first cam portion, a second cam portion is provided which increases the cam acceleration greatly and then gradually decreases to a negative acceleration, and further has a negative cam until the top dead center following the second cam portion. A third cam portion having a cam acceleration substantially equal to a limit acceleration at which the cam and the plunger do not separate from each other even at the maximum reciprocating frequency of the plunger is provided. In addition, in a suction stroke in which the plunger moves from a top dead center to a bottom dead center, a discharge stroke is performed. The cam profile has an acceleration curve ',', 'which is a target with respect to the cam acceleration and top dead center.

By lowering the negative cam acceleration over a wide range near the top dead center to the vicinity of the limit, the cam speed can be reduced in a short section, so that the peak position of the cam speed (= zero acceleration) can be easily set. Can be arranged in one third or more of the discharge stroke section. If the acceleration in the vicinity of the top dead center is set to a simple constant acceleration, a cam B shown in FIG. 4 is obtained, and the deceleration section becomes long because sufficient deceleration cannot be obtained. It becomes difficult to arrange.

Here, the limit acceleration indicated by the dotted line indicates a limit acceleration at which the spring force and the inertia force are balanced and no jumping occurs even at the maximum operation speed of the pump. The lower limit of the minimum acceleration in FIG. 3 is naturally determined from the allowable value of the Hertz stress of the cam and the limit of the size of the base circle, and the value of the spring force is determined in accordance with the lower limit. Is determined.

As described above, after the negative cam acceleration is reduced as much as possible, the cam acceleration from the bottom dead center is suppressed to a low value, and the rise of the cam speed is suppressed to a lower value by adopting a cam profile for re-acceleration halfway. Therefore, the maximum cam speed peak can be brought further backward. However, if the peak position is later than ２, the discharge amount is reduced, so that the acceleration value, the re-acceleration section and the like are appropriately adjusted to almost 1/3 of the discharge stroke section.
The peak of the cam speed comes to half. On this occasion,
Since the cam speed at the top dead center / bottom dead center is 0, the positive and negative areas of the cam acceleration must be the same.

Preferably, the cam acceleration at the bottom dead center is made continuous, that is, the cam acceleration (jerk) at the bottom dead center is set to zero. As a result, the change in the inertial force becomes continuous, so that high-order vibration components do not appear and low vibration and low noise can be achieved. Also, noise countermeasures become easier to carry out.

In the cam of FIG. 3 described above, in order to suppress the cam speed in the invalid section as much as possible, the acceleration curve is such that the starting acceleration is kept low and acceleration is resumed from the middle, but sufficient deceleration can be obtained. In such a case, such a re-acceleration is not performed, and a cam acceleration that monotonously decreases from the bottom dead center to the top dead center as shown in FIG. 7 may be used.

FIG. 7 is a cam showing another embodiment of the present invention.

In FIG. 7, following the first cam portion whose cam acceleration from the bottom dead center is set to a substantially constant positive value, the second cam portion which gradually reduces the cam acceleration and monotonously decreases to a negative acceleration. The negative cam acceleration to the top dead center following the second cam portion is set to a substantially constant negative value, and the negative acceleration is set to a limit acceleration at which the cam and the plunger do not separate at the maximum reciprocating frequency of the plunger. A third cam portion having substantially the same value is provided, and in a suction stroke in which the plunger moves from top dead center to bottom dead center, a cam profile having a target acceleration curve with respect to the cam acceleration in the discharge stroke and the top dead center. Things. Preferably, the jerk at the bottom dead center is set to 0 as in the case of the cam of FIG.

Although the rise of the cam speed is slightly larger than that of the cam of FIG. 3, practically sufficient discharge performance can be secured. Further, since the change in cam acceleration becomes smoother because re-acceleration is not performed, there is an advantage that vibration and noise are more advantageous. As shown in the figure, it is more practical to slightly increase the negative acceleration near the top dead center from the limit acceleration to provide a margin for jumping.

FIG. 8 is a view for explaining the cam shown in FIG. 3 and the cam shown in FIG. 7 according to one embodiment of the present invention.

In FIG. 8, the magnitude of the frequency component of the cam acceleration is shown. The horizontal axis is the vibration order, and the vertical axis is the power spectrum. Since there is no discontinuous portion in the acceleration waveform of both cams, the high-order vibration component is extremely small, and both can be a low vibration and low noise pump. Although the cam of FIG. 7 has a slightly lower volumetric efficiency than the cam of FIG. 3, the change in the inertial force proportional to the acceleration change is smoother.
It has the advantage of being superior in terms of vibration and noise because the higher-order vibration components are further reduced. Which cam to use depends on
It depends on whether weight is given to volumetric efficiency or vibration / noise, and it can be said that there is no difference between superiority and inferiority.

As shown in FIG. 1, the cam 100 according to an embodiment of the present invention has a common shaft with the engine cam shaft 72, and if the cam 100 is mounted in an additional form, the structure is simplified and the cost is reduced. There is an effect that can be achieved. In this example, the high-pressure fuel pump 101 is fixed to the engine cover 73 and the cam 100 is pressed against the lifter 3. However, the pump may be mounted on an engine block or the like, and the mounting direction may be upward, downward, lateral, or the like. It can be mounted in any orientation. The cam 100 may of course be provided inside the pump body 1, and the transmission of power from the engine camshaft may be performed via a coupling or the like.

As described above, according to the present invention, the high-pressure fuel pump 101 can obtain high volumetric efficiency over a wide range from the low rotation speed range to the high rotation speed range. In this case, sufficient fuel can be supplied at a high pressure, and the kinetic performance of a vehicle equipped with this engine is improved. Because of high volumetric efficiency,
The displacement of the pump can be reduced while the required discharge flow rate is secured, and the driving torque is reduced to improve fuel economy.

Further, since the absolute value of the negative cam acceleration near the top dead center can be minimized in the cam 100 of the present invention, the radius of curvature of the cam increases, the Hertz stress decreases, and the durability improves. . Alternatively, the base circle of the cam can be reduced while keeping the Hertz stress within an allowable value, so that the pump is small and easy to dispose in the engine room. At the same time, the driving torque is reduced, so that the fuel efficiency of the vehicle equipped with this pump is improved.

[0063]

According to the present invention, a high volumetric efficiency can be obtained in a wide range from a low rotation speed range to a high rotation speed range, and the size is small.
A high-pressure fuel pump having high torque efficiency and low vibration and noise can be provided.

[Brief description of the drawings]

FIG. 1 is a longitudinal sectional view illustrating the structure of a general high-pressure fuel pump.

FIG. 2 is an overall configuration diagram of a fuel supply device for an engine using a general high-pressure fuel pump.

FIG. 3 is a view showing an embodiment of a cam according to the present invention.

FIG. 4 is a diagram showing a difference in shape between a cam according to an embodiment of the present invention and another cam.

FIG. 5 is a diagram showing discharge flow characteristics of a high-pressure fuel pump according to an embodiment of the present invention.

FIG. 6 is a diagram showing a reason for a difference in flow rate characteristics depending on a cam shape.

FIG. 7 is a view showing a cam provided with another embodiment of the present invention.

FIG. 8 is a diagram illustrating frequency components of acceleration of a cam according to an embodiment of the present invention.

[Description of Signs] 1 ... pump body, 2 ... plunger, 3 ... lifter, 4 ... spring, 5 ... suction valve, 5a ... spring, 6 ... discharge valve, 6a ... spring, 10 ... suction flow path, 11 ... discharge flow Road, 12 ... Pressurizing chamber,
20: seal, 50: fuel tank, 51: low pressure pump,
52: low pressure regulator, 53: common rail, 54: injector, 55: high pressure regulator, 57: low pressure pipe, 71: engine, 72: engine camshaft, 73: engine cover, 100: cam, 1
01 ... High pressure fuel pump.

──────────────────────────────────────────────────の Continued on the front page (51) Int.Cl. ⁷ Identification symbol FI Theme coat ゛ (Reference) F04B 53/10 F04B 21/02 E 53/14 21/04 Z (72) Inventor Tadahiko Nogami Tsuchiura, Ibaraki 502, Kachimachi-cho, Machinery Research Laboratories, Hitachi, Ltd. F-term (Reference) 3G066 AA07 AB02 AC09 AD12 BA00 BA17 BA22 BA29 BA67 CA01S CA04T CA04U CA09 CA22T CE03 CE04 CE13 CE22 CE34 DB11 DB12 DB13 DC18 3H071 AA07 BB01 CC11 CC21 CC33 CC34 CC47 DD89 3H075 AA03 BB03 BB21 CC05 CC17 CC25 CC34 CC35 CC36 DA04 DA09 DB24 DB26

Claims (6)

[Claims] A pressurizing chamber communicating with a fuel intake passage and a discharge passage; a plunger reciprocating in the pressurized chamber by a shape of a rotating cam; a suction valve for opening and closing the suction passage; A high-pressure fuel pump having a discharge valve for opening and closing a passage, wherein while the suction valve is opened immediately after the plunger starts a discharge stroke from bottom dead center to top dead center, the rising speed of the plunger is suppressed. A high-pressure fuel pump having a shape of the cam for increasing the speed after the suction valve is closed. 2. A pressurizing chamber communicating with a fuel intake passage and a discharge passage, a plunger reciprocating in the pressurized chamber by a shape of a rotating cam, a suction valve for opening and closing the suction passage, and a discharge valve. A high-pressure fuel pump having a discharge valve that opens and closes a passage, wherein the high-pressure fuel pump includes an elastic member that presses the plunger in opposition to the urging force of the cam, and the plunger is disposed in a discharge stroke from a bottom dead center to a top dead center. Following the first cam portion, in which the cam acceleration from the bottom dead center is suppressed to a small positive value, a second cam portion is provided, in which the cam acceleration is largely increased and then gradually reduced to a negative acceleration. A third cam portion provided with a negative cam acceleration following the second cam portion up to the top dead center at substantially the same value as a limit acceleration at which the cam and the plunger do not separate from each other even at the maximum reciprocating frequency of the plunger. And a high-pressure fuel pump having a cam profile having a target acceleration curve with respect to a cam acceleration of the discharge stroke and a top dead center in a suction stroke of the plunger from a top dead center to a bottom dead center. . 3. A high-pressure fuel pump according to claim 2, wherein said cam acceleration has a cam profile that makes the cam acceleration at bottom dead center and top dead center substantially zero. 4. A pressurizing chamber communicating with a fuel suction passage and a discharge passage, a plunger reciprocating in the pressurizing chamber by a shape of a rotating cam, a suction valve for opening and closing the suction passage, and a discharge valve. A high-pressure fuel pump having a discharge valve for opening and closing a passage, comprising: an elastic member that presses the plunger in opposition to the urging force of the cam, wherein the cam discharges the plunger from a bottom dead center to a top dead center. Following the first cam portion, in which the cam acceleration from the bottom dead center in the stroke is a substantially constant positive value, a second cam portion is provided, in which the cam acceleration is gradually reduced to monotonically decrease to a negative acceleration. The negative cam acceleration to the top dead center following the second cam portion is set to a substantially constant negative value, and the negative acceleration is set to a limit acceleration at which the cam and the plunger do not separate from each other even at the maximum reciprocating frequency of the plunger. Ho A third cam portion having the same value is provided, and in a suction stroke in which the plunger is moved from a top dead center to a bottom dead center, a cam profile having a target acceleration curve with respect to the cam acceleration in the discharge stroke and an upper dead center is defined. A high-pressure fuel pump, comprising: 5. The high-pressure fuel pump according to claim 3, wherein the cam has a maximum peak position of the cam speed between approximately 1/3 and 1/2 of a discharge stroke section. 6. The high-pressure fuel pump according to claim 1, wherein the pump does not include an urging means such as an electromagnetic actuator capable of opening or closing the suction valve at an arbitrary timing.