JPH0353478B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION <Technical Field> The present invention relates to a vane pump suitable for a power steering device.

<Prior Art> Generally, a pressure balanced vane pump having 10 vanes is often used in a power steering device. This type of pump is lighter and easier to process than a vane pump with 12 vanes, but it tends to cause pressure pulsations due to fluctuations in the discharge flow rate. The following two main factors can be cited for generating pulsation in such a pump. One is the theoretical discharge flow rate variation determined by the cam dimensions and vane dimensions, and the other is the leakage flow rate variation due to changes in pump internal leakage, especially changes in the pump stage related to leakage.

Note that the flow rate fluctuation here is the difference between the maximum value and the minimum value of the flow rate waveform corresponding to the rotation angle θ of the vane, and the integral value of the flow rate waveform (absolute value of the flow rate)
This only affects pump efficiency and has nothing to do with pulsation.

Generally, as shown in the exploded view in Figure 1, the vane V
The cam curve with which it slides is made up of a suction curve section C1, a large circular section C2, a discharge curve section C3, and a small circular section C4. In this type of pump, when the vane V is moved by a unit angle Δθ, the change in volume surrounded by the vane at the final end and the vane at the front end of the pump chamber related to the output port OP is as follows. This is the discharge flow rate. This flow rate is large circle part C2, small circle part C
4 are perfectly circular, they will always be constant and will not fluctuate, but normally the large circular portion C2 is given a slight preload gradient due to preload, and therefore the discharge flow rate per unit angle will be equal to the preload gradient. As shown in FIG. 2, the flow fluctuation curve has a relatively small fluctuation width X1. This is generally referred to as basic discharge flow rate variation.

Furthermore, when the vane V related to the suction port IP moves by a unit angle, it rises due to the pressure action in the vane back pressure groove G leading to the air outlet port OP, so the air flow rate is consumed by the volume change on the lower end side of the vane. It turns out. Therefore, the consumed flow rate is proportional to the amount of rise of the vane V per unit angle, and corresponds to the velocity diagram of the vane locus (A in FIG. 1). For example, now, the suction curve part C1
Assuming that the vane V1 is located at the position α, the other vane V2 related to the suction curved portion c1 is located at the position α+
Exists at a position of 36°. These two vanes V1,
As the rotor R rotates, one of them is gradually increased in speed, and the other is gradually decelerated as the rotor R rotates. As the rotation progresses further, only one vane V1 exists on the suction curve C1, and this vane V
1 reaches the maximum speed and is then decelerated, the following speed fluctuations occur. Therefore, the consumption flow rate to the lower end of the vane in the suction area is the sum of the flow rates depending on the rising speed of one or two vanes.
It varies greatly depending on the angular position of the vane. However, the range of variation increases as the thickness of the vane increases.

From this, the theoretical discharge flow rate fluctuation determined by the cam specifications and vane specifications is the fluctuation width This theoretical discharge flow rate fluctuation is one of the causes of discharge pressure pulsation.

On the other hand, two adjacent vanes V, cam ring C,
The pressure in the plurality of pump chambers (sectors) divided by the rotor R and the side plate is determined by the vane V
The suction pressure and the discharge pressure change periodically depending on the angular position of the pressure. On the other hand, the vane back pressure groove G is always in a state of discharge pressure, and since there is a slight gap between the roller R and the side plate, the vane back pressure groove G Pressure oil leaks from the suction pressure to each sector. Moreover, in a pressure-balanced pump with 10 vanes, there is a problem in that the stage at which leakage occurs changes periodically. In other words, the discharge section is one vane section.
The data is alternately switched between 3 sectors and 2 sectors at a rate of 3 times. The solid line state in FIG. 1 indicates the starting state where the discharge section is 3 sectors, and the two-dot chain line state indicates the discharge section is 2 sectors.
It shows the starting state of becoming a sector. Since the section excluding this discharge section becomes the stage where leakage occurs, a difference is brought about in the stage (angular region). This causes leakage flow rate fluctuations as shown in FIG. 3B.

The pump discharge flow rate indicated by the solid line is the result obtained by subtracting the leakage flow rate fluctuation from the theoretical discharge flow rate fluctuation described above, although the theoretical discharge flow rate fluctuation is uniquely determined by the cam specifications and vane specifications. On the other hand, the leakage flow rate fluctuation is mainly determined as a function of the differential pressure between the vane back pressure groove and the suction section, and the fluctuation width X3 increases according to the load pressure. Therefore, when the pump is operated under no load, the differential pressure is small and the leakage flow rate fluctuation has little effect, so the actual discharge flow rate fluctuation is greatly influenced by the theoretical discharge flow rate fluctuation. However, when the differential pressure increases as the pressure increases, the leakage flow rate fluctuation exceeds the theoretical discharge flow rate fluctuation, and the actual discharge flow rate fluctuation becomes
It is greatly affected by fluctuations in leakage flow rate.

In particular, for vane pumps used in power steering devices, the load pressure changes significantly, so it is important to suppress fluctuations in the discharge flow rate due to pressure changes, and it is also a good idea to reduce fluctuations in the theoretical output flow rate itself. .

<Object of the Invention> Therefore, an object of the present invention is to provide a vane pump that does not change the pump stage related to leakage in order to suppress fluctuations in leakage flow rate, thereby reducing fluctuations in discharge flow rate due to the influence of load pressure. The goal is to reduce pressure pulsations.

Another object of the present invention is to suppress fluctuations in the theoretical discharge flow rate in addition to the fluctuations in the leakage flow rate described above, thereby making fluctuations in the pump discharge flow rate extremely small and significantly reducing pressure pulsations.

<Examples> Examples of the present invention will be described below based on the drawings. 4 and 5, reference numeral 10 indicates a pump housing, and a hollow chamber 11 with a bottom is formed in this pump housing 10.
is open at one end of the pump housing 10.
An end cover 12 is fixed to one end of the pump housing 10 to close the opening thereof. Inside the hollow chamber 11 surrounded by the pump housing 10 and the end cover 12, there is a cam ring 14, a ring-shaped side plate 15 that faces one side of the cam ring 14, and a ring-shaped side plate 15 that has one end that faces the other side of the cam ring 14. A disk-shaped side plate 16 whose back surface is in contact with the end cover 12 is housed, and one side plate 15 is connected to the pump housing 10.
is fitted into the bearing hole. A web washer 17 is inserted between one side plate 15 and the pump housing 10 in a resilient state, and the force of the web washer 17 causes the cam ring 14, the pair of side plates 15 and 16, and the end cover 10 to are in contact with each other. In addition,
Cam ring 14 and a pair of side plates 1
5 and 16 are a pair of positioning pins 18 supported between the pump housing 10 and the end cover 12.
The phase is determined by

A substantially elliptical cam surface 20 (described later) is formed on the inner periphery of the cam ring 14, and a rotor 22 in which ten vanes 21 that are in sliding contact with the cam surface 20 are fitted so as to be slidable in the radial direction is mounted inside the cam ring 14. It is stored in. Such rotor 22 and vane 21
The axial width of the cam ring 14 is determined to be somewhat smaller than the axial width of the cam ring 14, and when the pair of side plates 15, 16 are in contact with both side surfaces of the cam ring 14, the rotor 22 and the side plates 15, 16 are Appropriate side clearance (axial clearance) is maintained between the two. The rotor 22 is splined to one end of a rotating shaft 24 rotatably supported by a bearing sleeve 23 fitted into a bearing hole of the pump housing 10 .

With the above configuration, the cam surface 2 of the cam ring 14
A plurality of pump chambers partitioned by vanes 21 are formed between the rotor 22 and the outer peripheral surface of the rotor 22, and the volume of each pump chamber changes as the rotor 22 rotates. On each side of the pair of side plates 15, 16 that are in contact with the rotor 22, suction ports 25, 26 are provided corresponding to the pump chambers that perform the expansion process, and discharge ports 2 are provided that correspond to the pump chambers that perform the compression process.
7 and 28 are each formed at two locations on the circumference. The suction ports 25 and 26 are provided in a suction chamber 29 recessed in the hollow chamber 11 so as to surround the cam ring 14.
This suction chamber 29 is opened to a suction passage 31 that communicates with a reservoir 30 and a bypass passage 33 that communicates with a flow rate regulating valve 32 . One discharge port 27 passes through the side plate 15 and communicates with a discharge chamber 34 formed between the side plate 15 and the pump housing 10 . It is communicated with the pressure fluid outlet via the throttle passage, and is also appropriately communicated with the bypass passage 33 via the flow rate regulating valve 32. Also integrated side plates 15,1
6, each surface facing the rotor 21 is provided with an annular or arcuate vane back pressure groove 3 corresponding to the lower end of the vane.
7 and 38 are formed, and one vane backpressure groove 37 is communicated with the discharge chamber 34 through a through hole 39 to introduce discharge fluid to the lower end of the vane.

Next, the cam surface shape of the cam ring 14, the suction ports 25, 26, the discharge ports 27, 2
The specific configuration of No. 8 will be explained.

FIG. 6 shows a developed view of the pump in a half-circumference portion from 0 to 180 degrees, and the remaining half-circumferences are constructed in the same manner. The cam surface 20 consists of a cam curve that smoothly connects a constant acceleration suction curve section C1, a large circular section C2 with a slight preload gradient, a discharge curve section C3, and a small circular section C4. Specifically, the suction section (suction curve section C1) is set to be considerably longer than the conventional one, and the discharge section (discharge curve section C3) is set to be considerably shorter.

The starting ends of the suction ports 25 and 26 opened corresponding to the suction curved portion C1 and the terminal ends of the protruding ports 27 and 28 opened corresponding to the protruding curved portion C3 are as follows.
When the center of the vane 21 corresponds to the starting end of the suction ports 25, 26 as shown in FIG. It is determined in the relative position corresponding to. However, suction port 2
The angular width of 5 and 26 is 2 of the angular pitch of the vane 21.
The angular width of the discharge ports 27 and 28 is determined to be the angular pitch of the vane 21 plus an angle corresponding to the thickness of the vane 21.

From this, if one vane 21 is located at the starting end of the suction ports 25, 26, the preceding first vane 21 is located at the starting end of the suction ports 25, 26.
, the second vane 21 is located at a position slightly deviated from the terminal ends of the suction ports 25 and 26,
The third vane 21 is located at a position in contact with the starting ends of the discharge ports 27 and 28, and the fourth vane 21 is located at a position connected to the terminal ends of the discharge ports 27 and 28, respectively.

Since the vane pump of the present invention is configured as described above, when the rotating shaft 24 is driven, the rotor 22 rotates, thereby causing the working fluid to flow into the suction chamber.
9 into the pump chamber through the suction ports 25 and 26, and discharge fluid is discharged from the pump chamber through the discharge ports 27 and 28 into the discharge chamber 34. The flow rate is controlled to a predetermined flow rate by a flow rate regulating valve 32, and then sent to a power steering device or the like.

When the discharge pressure of the pump increases, the vane back pressure grooves 37 and 38 on the high-pressure side pass through the side clearance between the integrated side plates 15 and 16 and the rotor 22
Pressure fluid leaks from the suction ports 25 and 26 on the low pressure side.

However, in the present invention, by setting the angular width of the discharge ports 27 and 28 to the angular pitch of the vane 21 plus an angle corresponding to the thickness of the vane 21, suction is drawn from the vane back pressure grooves 37 and 38. The number of sectors (pump chambers) in which leakage occurs toward the ports 25 and 26 can be maintained at a constant angular width (two pump chambers) at any angular position of the vane 21. This can be easily understood by considering the situations before and after the situation shown in FIG.

In a state slightly before that shown in FIG. 6, as shown in FIG. Room Ps, 4th, 5th
and the first three vanes 21D, 21E, 21
The two pump chambers defined by A have a discharge pressure Pd.
and two third and fourth vanes 21C,
The pump chamber defined by 21d has a suction pressure Ps
The pressure Pm is higher than the discharge pressure Pd and lower than the discharge pressure Pd. Therefore, leakage of pressure oil from the vane back pressure grooves 37, 38 toward the suction ports 25, 26 occurs in the angular region θ1 defined by the four first to fourth vanes 21A to 21D.

From the state shown in Fig. 7, the state shown in Fig. 6 is passed and the state shown in Fig. 8 is reached.
When the state shown in the figure is reached, the pump chamber defined by the third and fourth vanes 21C and 21D becomes completely at the discharge pressure Pd, and the pump chamber defined by the fifth and first vanes 21E and 21A reaches the discharge pressure Pd. The pump chamber switches from the discharge pressure Pd to the suction pressure Ps. However, in this case as well, the vane back pressure groove 3
Pressure oil leaks from 7, 38 toward suction ports 25, 26.

In this way, when the discharge area changes from the state shown in FIG. 7 to the state shown in FIG. 5 as the vane 21 moves, and further from the state shown in FIG. 5 to the state shown in FIG. 8, in the state shown in FIG. The number of pump chambers in the discharge area becomes one (or three) momentarily, but otherwise the number of pump chambers in the discharge area can always be reduced to two, and the angular width related to leakage is always kept constant. become able to. As a result, vane 21
No change in the stage related to leakage occurs at any angular position of the pump, and fluctuations in pump internal leakage can be made extremely small.

Moreover, with the above-described configuration, fluctuations in the flow rate consumed due to the rise of the vane 21 in the suction section can also be reduced.

That is, as shown in the velocity diagram A of FIG. 6, if the vane 21 is located at the starting end position α of the suction curve C1, the preceding adjacent vane 21 is at α+
It exists near the maximum speed position of 36 degrees, and as the rotor 22 rotates, one of the two vanes 21 gradually increases in speed, and the other gradually decelerates. When the vane 21 on the speed increasing side reaches the maximum speed position,
The deceleration side vane 21 reaches almost the terminal position of the suction curved portion C1, and at the same time, the succeeding vane 21 reaches the starting end position of the suction curved portion C1. The total amount of flow consumed due to the rise of the vanes 21 in the suction section is
1, but since there are always two vanes 21 in the suction section, the consumption flow rate is always constant, and as a result, the theoretical discharge flow rate fluctuation is It is not affected by the thickness of the vane 21, and the third
Only the flow rate fluctuation shown in the figure will occur, and the fluctuation range will be extremely small.

<Effects of the Invention> As described above, the present invention has a configuration in which the stage related to leakage between the vane back pressure groove and each pump chamber does not change depending on the rotational position of the vane, so that the internal The range of flow rate fluctuations due to leakage can be made extremely small, thereby suppressing fluctuations in discharge flow rate and reducing pressure pulsations.

Moreover, the present invention is configured to reduce fluctuations in leakage flow rate as described above, prevent fluctuations in flow consumption due to the rise of the vane in the suction section, and reduce fluctuations in theoretical discharge flow rate, so fluctuations in flow rate can be effectively reduced. The effect of significantly reducing pressure pulsation can be achieved.

[Brief explanation of drawings]

Figure 1 is a developed diagram showing an example of a conventional vane pump, Figure 2 is a diagram showing fluctuations in basic discharge flow rate, Figure 3 is a diagram showing fluctuations in theoretical discharge flow rate and leakage flow rate, and Figure 4 is a diagram showing variations in basic discharge flow rate. A sectional view of a vane pump showing an embodiment of the invention, FIG. 5 is a sectional view taken along the line V-V in FIG. 4, and FIG. 6 is an expanded view of a part of FIG.
7 and 8 are diagrams showing the operating state of FIG. 6, respectively. 10... Pump housing, 12... End cover, 14... Cam ring, 15, 16... Side plate, 20... Cam surface, 21... Vane,
22... Rotor, 25, 26... Suction port, 2
7, 28...Discharge port, 37, 38...Vane back pressure groove.

Claims (1)

[Scope of Claims] 1. A pump housing, a cam ring housed in the pump housing, a rotatable rotor having 10 vanes in sliding contact with the cam ring held at equal angular intervals on the circumference, and arranged on the side of the cam ring. A pressure-balanced vane pump comprising a side plate having a circumferentially formed suction port, a discharge port, and a vane back pressure groove communicating with the discharge port, the pump having a plurality of pump chambers partitioned by the vanes. The discharge port is arranged at an angle formed by each outer end of two adjacent vanes so that the stage related to leakage caused by a pressure difference between the vane and the back pressure groove does not change at any angular position of the vane. A vane pump that is set approximately equal to the width. 2. A pump housing, a cam ring that is housed in the pump housing and has a cam portion on its inner periphery that has a constant acceleration suction curve section and a discharge curve section, and 10 vanes that slide into contact with this cam ring are arranged at equal angular intervals on the circumference. A pressure-balanced vane pump comprising a rotatable rotor held, a side plate disposed on the side of the cam ring and having a suction port, a discharge port, and a vane back pressure groove communicating with the discharge port formed in a circumferential shape, The discharge port is configured such that the stage related to leakage caused by a pressure difference between each of the plurality of pump chambers partitioned by the vane and the vane back pressure groove does not change at any angular position of the vane. , is set to be approximately equal to the angular width formed by each outer end of two adjacent vanes, and the angular pitch of the vanes is set so that the angle formed by the suction curved portion is always such that two vanes are present in this suction curved portion. A vane pump that is set at almost twice the speed.