Oil pump pressure control device

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to an oil pump pressure control device for a variable flow rate oil pump in which two discharge ports are provided and which uses three rotors as means for provision of the two discharge sources thereof that facilitates reduced friction while maintaining pressure (oil pressure) characteristics not in multi stage characteristics but in characteristics identical to the pressure characteristics of a common oil pump based on a newly devised method of switching oil passages.

2. Description of the Related Art

As disclosed in Japanese Unexamined Patent Application No. H11-280666, the use of three rotors that themselves serve as means for transporting oil to two locations is well known in the conventional art. Japanese Unexamined Patent Application No. H11-280666 refers only to means for transporting oil to two locations and makes no reference to the subsequent effects thereof.

SUMMARY OF THE INVENTION

As described in Japanese Unexamined Patent Application No. H11-280666, the tooth profile is determined once the flow rate produced by an inner side rotor and outer side rotor is established and, accordingly, the flow rate ratio remains substantially constant no matter what the revolutions. More particularly, a pressure difference is generated between the space between the inner-side teeth and the space between the outer-side teeth due to the characteristics of the inner-side tooth profile and the outer-side tooth profile which, as a result, there is a fear of the middle rotor being pressed against one side and, in turn, of the wear of the teeth surface thereof being accelerated. In other words, a significant drawback inherent to this device pertains to the generation of biased wear on the middle rotor caused by pressure difference.

Japanese Unexamined Patent Application No. 2005-140022 describes a device designed with the aim of decreasing superfluous work and increasing efficiency at the low revolution range based on oil being relieved (returned) at a desired revolution range. Referring to FIG. 8 of page 13 thereof, superfluous work is reduced and efficiency is increased as a result of drop in the flow rate at a desired number of revolutions. While the efficiency can be increased by the adoption of this kind of variable flow rate oil pump, the following problems are inherent thereto.

That is to say, there is a drawback inherent to this device in that, because the discharge port of the single rotor assembly thereof is partitioned into two, the opened area of the respective discharge passages thereof is reduced and the rotor diameter must be increased to compensate for this reduction, and this increase in rotor diameter results in increased friction (torque) and a drop in efficiency. In addition, despite the aim of the device being increased efficiency based on a variable capacity specification for eliminating superfluous work, an undesirable drop in efficiency occurs. Furthermore, because the discharge port of the single rotor assembly is simply partitioned into two, the rotor characteristics of the two discharge ports are of course equivalent and, accordingly, there is a limit to the degree of freedom available in terms of the distribution ratio of the two discharge amounts. In other words, the port position and opened area are governed by fixed design restrictions. There is an additional drawback in that, because the two discharge sources constitute a discharge from the same rotor, overlapping of the discharge pulse timing occurs and an overlapping and magnification of the noise and vibration thereof occurs. Another drawbacks pertains to the rotor being a single rotor assembly and, accordingly, to the fact that there is no means available for reducing superfluous work when it occurs.

Thereupon, the problem to be solved by the present invention (technical problem and object and so on) is to produce a three rotor-using oil pump for a variable flow rate oil pump in which two discharge ports are provided and which uses three rotors as means for provision of the two discharge sources thereof that facilitates reduced friction while maintaining pressure (oil pressure) characteristics identical to the pressure characteristics of a common oil pump (The oil pump of Japanese Unexamined Patent Application No. 2002-70756 that exhibits the non-linear stepped characteristic passing through the broken line as shown in FIG. 10 of page 7 thereof, and comprises a valve with a ON/OFF relief function. In addition, which exhibits approximately one characteristic inflection point.) rather than exhibiting multi-stepped characteristics based on a newly devised method of switching oil passages.

Thereupon, as a result of exhaustive research conducted by the inventors with a view to resolving the problems described above, the abovementioned problems were able to be solved by the oil pump pressure control device of the invention of claim 1 comprising: an oil pump having three rotors including an outer circumferential side rotor and an inner circumferential side rotor; a first discharge passage for feeding oil from the abovementioned outer circumferential side rotor to an engine; a return passage that returns to an intake side of the abovementioned outer circumferential side rotor; a second discharge passage for feeding oil from the abovementioned inner circumferential side rotor to the engine; a return passage that returns to an intake side of the abovementioned inner circumferential side rotor; and a pressure control valve whose valve main body is provided between the discharge port from the abovementioned inner circumferential side rotor and the abovementioned first discharge passage, wherein the abovementioned first discharge passage and the abovementioned second discharge passage are coupled, and a flow passage control is executed in each of: a low revolution range in a state in which only the first discharge passage and the second discharge passage are open; in an intermediate revolution range in a state in which the first discharge passage and the second discharge passage are open and a first inlet side of the abovementioned return passage is closed while a second inlet side of the abovementioned second return passage is open; and a high revolution range in a state in which the second discharge passage is closed while the first discharge passage and the abovementioned return passage are open.

The abovementioned problems were able to be solved by the oil pump pressure control device of the invention of claim 2 comprising: an oil pump having three rotors including an outer circumferential side rotor and an inner circumferential side rotor; a first discharge passage for feeding oil from the abovementioned outer circumferential side rotor to an engine; a first return passage that returns to an intake side of the abovementioned outer circumferential side rotor; a second discharge passage for feeding oil from the abovementioned inner circumferential side rotor to the engine; a second return passage that returns to an intake side of the abovementioned inner circumferential side rotor; and a pressure control valve whose valve main body is provided between a discharge port from the abovementioned inner circumferential side rotor and the abovementioned first discharge passage, wherein the abovementioned first discharge passage and the abovementioned second discharge passage are coupled, and a flow passage control is executed in each of: a low revolution range in a state in which only the first discharge passage and the second discharge passage are open; an intermediate revolution range in a state in which the first discharge passage and the second discharge passage are open and the abovementioned first return passage is closed while the abovementioned second return passage is open; and a high revolution range in a state in which the second discharge passage is closed while the first discharge passage is open and the abovementioned first return passage and the abovementioned second return passage are open.

The abovementioned problems are able to be resolved by the oil pump pressure control device of the invention of claim 2 according to the configuration described above in which the abovementioned pressure control valve is configured as a type comprising a first valve portion and a second valve portion. The abovementioned problems are able to be resolved by the oil pump pressure control device of the invention of claim 3 according to the configuration described above in which the abovementioned pressure control valve is configured as a type comprising a first valve portion, a second valve portion and a third valve portion.

A significant merit of the invention of claim 1 is that, even in variable flow rate oil pumps that use three rotors, a pressure ratio determined at the design stage can eliminate the other pressure difference produced by the discharge passages and return passages and can prevent a middle rotor from which the inner circumferential side rotor is configured from being pressed to one side whereupon, in turn, friction on the teeth surface can be prevented and the durability thereof improved. A further effect thereof is that at times of high-speed revolution of the outer circumferential side rotor and the inner circumferential side rotor the second discharge passage of the inner circumferential side rotor is fully closed so as to form the inner circumferential side rotor as an independent circuit whereupon, even in the absence of the generation of a superfluous work pressure by the inner circumferential side rotor, there is no drop in overall pump pressure. In addition, because work=pressure×flow rate, the superfluous work can be reduced so long as the pressure is lowered and, because the outer circumferential side rotor pump (main pump) and the inner circumferential side rotor pump (sub pump) are not in communication, the sub pump pressure is significantly lowered.

In addition, because the inner circumferential side pump (sub pump) forms an independent circuit at times of high revolution, if the opened area return passage of the pump is increased more oil can be discharged and the pump pressure further reduced. Because a single discharge port is not partitioned into two for two discharge sources thereof and instead a single discharge port can be established for each, the flow rate is not partitioned. Accordingly, because the rotor diameter of the three rotors can be reduced and the slide area of the rotors decreased comparative to a common single rotor assembly, the friction (torque) can be reduced and the pump efficiency increased. An additional merit of the invention is that, because the three rotors can be configured as a rotor having a dual-rotor assembly, the superfluous work of one rotor can be reduced. The invention of claim 2 affords an identical effect to the effect afforded by the invention of claim 1.

In addition, as well as an effect identical to the effect afforded by the invention of claim 1, the invention of claim 3, based on the provision of a simple pressure control valve, affords a reduction in the number of component parts. In addition, the invention of claim 4, based on the provision of a 3-valve-type pressure control valve, facilitates a pressure control by which desired values can be better approximated.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a systems diagram of a first embodiment of the present invention showing a state in the engine low revolution range;

FIG. 2 is a systems diagram of the first embodiment of the present invention showing a state in the engine intermediate revolution range;

FIG. 3 is a systems diagram of the first embodiment of the present invention showing a state in the engine high revolution range;

FIG. 4 is a systems diagram of a second embodiment of the present invention showing a state in the engine low revolution range;

FIG. 5 is a systems diagram of part of the second embodiment of the present invention showing a state in the engine intermediate revolution range;

FIG. 6 is a systems diagram of a part of the second embodiment of the present invention showing a state in the engine high revolution range;

FIG. 7 is a simplified systems diagram of the present invention; and

FIG. 8A is a characteristics graph of the engine revolutions and discharge pressure of the present invention, and FIG. 8B is a characteristics chart of the engine revolutions and discharge flow rate of the present invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

In a description of the embodiments of the present invention given hereinafter with reference to the drawings, the device shown in the diagrams is a three rotor-type oil pump principally configured from an outer rotor 1, a middle rotor 2 and an inner rotor 3. An outer circumferential side intake port 4 and an outer circumferential side discharge port 5 are provided in the abovementioned outer rotor 1 and the abovementioned middle rotor 2, and an inner circumferential side intake port 6 and inner circumferential side discharge port 7 are provided in the abovementioned middle rotor 2 and inner rotor 3. The abovementioned outer rotor 1, the abovementioned middle rotor 2, the outer circumferential side intake port 4 and the outer circumferential side discharge port 5 generically describe an outer circumferential side rotor, and the abovementioned middle rotor 2, the abovementioned inner rotor 3, the inner circumferential side intake port 6 and inner circumferential side intake port 7 generically describe an inner circumferential side rotor.

The three rotor-type oil pump is configured from a first discharge passage 11 for feeding oil from the abovementioned outer circumferential side discharge port 5 to an engine E, a first return passage 12 that returns to an intake passage 8 of the outer circumferential side intake port 4, a second discharge passage 13 for feeding oil from the abovementioned inner circumferential side discharge port 7 to the engine E, and a second return passage 14 that returns to an intake passage 9 of the abovementioned inner circumferential side intake port 6, an end portion side of the abovementioned second discharge passage 13 being coupled with the abovementioned first discharge passage 11 at a suitable position therealong. In addition, the abovementioned intake passage 8 and the abovementioned intake passage 9 may generically describe an intake body D (see FIG. 4). In addition, the abovementioned first return passage 12 and the abovementioned second return passage 14 may generically describe a return passage E (see FIG. 4).

The symbol C denotes a pressure control valve configured from a valve main body 20 and a valve housing 30 which is provided across the abovementioned first discharge passage 11, first return passage 12, second discharge passage 13 and second return passage 14. The abovementioned valve main body 20 is configured from a first valve portion 21, a narrow-diameter coupling portion 23 and a second valve portion 22. A valve comprising the abovementioned first valve portion 21 and second valve portion 22 is referred to as a two-valve pressure control valve C. In addition, a long-hole portion 31 slidable in accordance with need with respect to the abovementioned valve main body 20 is formed in the pressure control valve C, the abovementioned valve main body 20 being constantly push-pressured from a cover body 33 fixed in a rear portion side of the second valve portion 22 to the abovementioned first valve portion 21 side by the elastic pressure produced by a compression coil spring 40 within the long-hole portion 31. The symbol 32 denotes a stopper portion formed in one end of the abovementioned long-hole portion 31 and positioned in a suitable position of the abovementioned first discharge passage 11.

In addition to the items that variously determine the pressure conditions, the diameter of the valve main body 20 and the spring constant of the compression coil spring 40, the control of the pressure control valve C also requires that various conditions dependent on change in the discharge pressure of the abovementioned first discharge passage 11 be satisfied. More specifically, a flow rate control must be executed in each of a low revolution range which constitutes a state in which only the first discharge passage 11 and the second discharge passage 13 are opened as shown in FIG. 1, an intermediate revolution range which constitutes a state in which the first discharge passage 11 and the second discharge passage 13 are opened and the abovementioned first return passage 12 is closed to open the second return passage 14 as shown in FIG. 2 and, in addition, in a high revolution range which constitutes a state in which the second discharge passage 13 is closed to open the first discharge passage 11 and the first return passage 12 and the second return passage 14 are open as shown in FIG. 3.

The operation of the pressure control valve C will be hereinafter described. First, in the low revolution range of the outer circumferential side rotor A and the inner circumferential side rotor B, in other words, when the engine revolutions are in the low revolution range which constitutes the state of FIG. 1, each of the return passages of the outer circumferential side rotor A and the inner circumferential side rotor B are closed by the first valve portion 21 and the second valve portion 22 of the pressure control valve C, and all oil discharged from the first discharge passage 11 and the second discharge passage 13 is discharged to the engine. The first discharge passage 11 of the outer circumferential side rotor A and the second discharge passage 13 of the inner circumferential side rotor B are in communication and, accordingly, an equalization of pressure occurs. In addition, because the return passages are closed, the overall discharge flow rate of the oil pump is equivalent to a sum of the flow rates of the outer circumferential side rotor A and the inner circumferential side rotor B. The characteristics produced in the low revolution range are shown in a characteristics graph of number of revolutions and discharge pressure (see FIG. 8A) and a characteristics graph of number of revolutions and discharge flow rate (see FIG. 8B).

A state in which the engine revolutions have risen further is taken as the intermediate revolution range. In this state, which constitutes the state of FIG. 2, an opening portion 141 of the second return passage 14 starts to open, and an opening portion 131 of the second discharge passage 13 starts to close. A more specific description thereof will be hereinafter provided. The first discharge passage 11 of the outer circumferential side rotor A and the second discharge passage 13 of the inner circumferential side rotor B remain in communication. Because the opening portion 141 of the second return passage 14 of the inner circumferential side rotor B starts to open, first, the rise in pressure in the inner circumferential side rotor B stops. Simultaneously, because the first discharge passage 11 and the second discharge passage 13 are in communication, a backflow of oil from the discharge of the outer circumferential side rotor A occurs and, in this state, is exhausted through the second return passage 14 of the inner circumferential side rotor B and returned to the intake passage 9 of the inner circumferential side rotor B. The state afforded by this series of actions results in a substantial equalization of the pressure of the outer circumferential side rotor A and the pressure of the inner circumferential side rotor B.

Because the opening portion 131 of the second discharge passage 13 of the inner circumferential side rotor B gradually closes and the opening portion 141 of the second return passage 14 of the inner circumferential side rotor B gradually opens consequent to a rise in the number of revolutions in the intermediate revolution range, the overall increase in the flow rate when the number of revolutions rises is negligible. In reality, the pressure not expressed in the true surface of the discharge of the inner circumferential side rotor B gradually drops due to the opening portion 141 of the second return passage 14 of the inner circumferential side rotor B being gradually opened. However, because the first discharge passage 11 and the second discharge passage 13 are in communication, an equalization of the pressure of the outer circumferential side rotor A and the inner circumferential side rotor B occurs, and the pressure of the inner circumferential side rotor B has the appearance of not dropping.

In addition, because the opening portion 121 of the first return passage 12 is still not open at the intermediate revolution range, the discharge flow rate of the outer circumferential side rotor A increases together with the number of revolutions. The discharge flow rate of the inner circumferential side rotor B decreases together with the number of revolutions and the opening portion 141 of the second return passage 14 of the inner circumferential side rotor B being opened. The backflow rate from the discharge of the outer circumferential side rotor A exceeds the discharge flow rate of the inner circumferential side rotor B subsequent to a certain number of revolutions being attained and, accordingly, the resultant discharge flow rate of the inner circumferential side rotor B is negative. The generation of a negative flow rate in this way means that a flow rate equivalent to a sum of the flow rate of two oil pumps can be produced and that a flow rate equivalent to less than a flow rate of a single pump can be produced and, accordingly, that a broad variation in flow rate is possible. The characteristics at the intermediate revolution range are expressed in the pressure characteristics graphs of revolutions with respect to discharge pressure and discharge flow rate (see FIG. 8) and, while the increase in the outer circumferential side rotor A is steady, a negative discharge flow rate is produced at the inner circumferential side rotor B side due to backflow, and a pressure linking line obtained as a sum of the outer circumferential side rotor A and inner circumferential side rotor B is substantially identical to the pressure characteristics of a conventional oil pump.

A state in which the engine revolutions have increased further is taken as the high revolution range. In this state, which is the state shown in FIG. 3, the opening portion 121 of the first return passage 12 starts to open and the opening portion 131 of the second discharge passage 13 has finished closing. A more specific description thereof will be hereinafter provided. Because the discharge of the inner circumferential side rotor B is fully closed, the discharge of the outer circumferential side rotor A and the discharge of the inner circumferential side rotor B are no longer in communication. That is to say, the inner circumferential side rotor B is formed as an oil circuit independent of the outer circumferential side rotor A. The pressure from the discharge of the outer circumferential side rotor A is unable to reach the inner circumferential side rotor B and instead is simply returned through the second return passage 14 of the inner circumferential side rotor B, and this results in an instant drop in the pressure of the inner circumferential side rotor B. Because backflow to the inner circumferential side rotor B also stops and all the oil discharged from the inner circumferential side rotor B is returned by way of the second return passage 14, a zero flow rate from the inner circumferential side rotor B to the engine E is established.

In other words, because the friction (torque) can be caused to drop instantly and superfluous work can be eliminated due to the zero flow rate from the inner circumferential side rotor B and the discharge of the inner circumferential side rotor B performing no work at all, the overall efficiency of the pump is increased. The characteristics at the high revolution range are expressed in the pressure characteristics graphs of revolutions with respect to discharge pressure and discharge flow rate (see FIG. 8) and, while there is a gradual increase in the outer circumferential side rotor A, the inner circumferential side rotor B is in a closed state and a pressure linking line obtained as a sum of the outer circumferential side rotor A and inner circumferential side rotor B is equivalent to the outer circumferential side rotor A alone. Because of the decrease in friction (torque) due to the drop in the pressure of the inner circumferential side rotor B in this way, the efficiency is increased.

Regarding the outer circumferential side rotor A pressure, while a return of oil occurs by way of the second return passage 14 in the intermediate revolution range because the first discharge passage 11 and the second discharge passage 13 are in communication, because of the continuous return of oil from the first return passage 12 that occurs in the high revolution range, the change in the outer circumferential side rotor pressure between the intermediate revolution range and the high revolution range is negligible. In addition, because the opening portion 121 of the first return passage 12 opens and overflow to the first return passage 12 occurs at the instant of opening thereof, the change in the outer circumferential side rotor A flow rate occurring subsequent to this drop in flow rate is negligible. Strictly speaking, very little rise occurs consequent to the rise in the number of revolutions.

Because the opening portion 131 of the second discharge passage 13 of the inner circumferential side rotor B is fully closed, the “pressure” of the pump main body (sum of the outer circumferential side rotor A and inner circumferential side rotor B) is equivalent to the pressure of the outer circumferential side rotor A alone. While the change in the pressure of the outer circumferential side rotor A is negligible due to the opening portion 121 of the first return passage 12 being open, strictly speaking, only very small increases in pressure occur consequent to increases in the number of revolutions. In addition, for the “flow rate” of the pump main body, because the opening portion 131 of the second discharge passage 13 of the inner circumferential side rotor B is fully closed, the “flow rate” of the outer circumferential side rotor A constitutes the overall pump flow rate. While the change in the pressure of the outer circumferential side rotor A is negligible due to the opening portion 121 of the first return passage 12 being open, strictly speaking, only very small rises in pressure occur consequent to increases in the number of revolutions.

Another embodiment of the pressure control valve C will be hereinafter described. The pressure control valve C is configured from a valve body 20 and valve housing 30 and is provided between the abovementioned first discharge passage 11, the first return passage 12, the second discharge passage 13 and the second return passage 14. The abovementioned valve main body 20 is configured from a first valve portion 21, a narrow-diameter coupling portion 23, a second valve portion 22, a third valve portion 24 and a narrow-diameter coupling portion 25. The remaining configuration thereof is identical to the configuration shown in FIG. 1 to FIG. 3. A valve comprising the abovementioned first valve portion 21, second valve portion 22 and third valve portion 24 is referred to as a three valve-type pressure control valve C.

The action thereof will be hereinafter described. First, in the low revolution range of the outer circumferential side rotor A and inner circumferential side rotor B, in other words, when the engine revolutions are in the low revolution range which constitutes the state of FIG. 4, the return passages of each of the outer circumferential side rotor A and the inner circumferential side rotor B are closed by the first valve portion 21 and the third valve portion 24 of the pressure control valve C, and all the oil discharged from the first discharge passage 11 and the second discharge passage 13 is discharged to the engine. Because the first discharge passage 11 of the outer circumferential side rotor A and the second discharge passage 13 of the inner circumferential side rotor B are in communication, an equalization of pressure occurs. In addition, because the return passages are closed, the discharge flow rate of the oil pump as a whole constitutes a sum of the flow rates of the outer circumferential side rotor A and the inner circumferential side rotor B.

A state in which the engine revolutions have risen further is taken as the intermediate revolution range. In this state, which constitutes the state of FIG. 5, the opening portion 141 of the second return passage 14 starts to open and the opening portion 131 of the second discharge passage 13 starts to close. A description thereof has been omitted. A state resulting from further increase in the engine revolutions is taken as the high revolution range. In this state, which constitutes the state of FIG. 6, the opening portion 121 of the first return passage 12 starts to open and the opening portion 131 of the second discharge passage 13 has finished closing. Because the discharge of the inner circumferential side rotor B is fully closed, the discharge of the outer circumferential side rotor A and the discharge of the inner circumferential side rotor B are no longer in communication. That is to say, the inner circumferential side rotor B forms an oil circuit independent of the outer circumferential side rotor A. The pressure from the discharge of the outer circumferential side rotor A is unable to reach the inner circumferential side rotor B and is instead simply returned through the second return passage 14 of the inner circumferential side rotor B resulting in an instant drop in the pressure of the outer circumferential side rotor B. Because backflow to the inner circumferential side rotor B also stops and all the oil discharged from the inner circumferential side rotor B is returned by way of the second return passage 14, a zero flow rate from the inner circumferential side rotor B to the engine E is established.

In other words, because the friction (torque) can be caused to drop instantly and superfluous work eliminated due to the zero flow rate of the inner circumferential side rotor B and the discharge of the inner circumferential side rotor B performing no work at all, the overall efficiency of the pump is increased. The characteristics at this high revolution range are expressed in the pressure characteristics graphs of revolutions with respect to discharge pressure and discharge flow rate (see FIG. 8) and, while the increase in the outer circumferential side rotor A is gradual, the inner circumferential side rotor B is in a closed state and a pressure linking line obtained as a sum of the outer circumferential side rotor A and inner circumferential side rotor B is equivalent to the outer circumferential side rotor A alone. Because of the decrease in friction (torque) due to the drop in the pressure of the inner circumferential side rotor B in this way, the efficiency is increased.

Regarding the outer circumferential side rotor A pressure, while a return of oil occurs by way of the second return passage 14 in the intermediate revolution range because the first discharge passage 11 and the second discharge passage 13 are in communication, because of the continuous return from the first return passage 12 that occurs in the high revolution range, the change in the outer circumferential side rotor pressure between the intermediate revolution range and the high revolution range is negligible. In addition, because the opening portion 121 of the first return passage 12 opens and overflow to the first return passage 12 occurs at the instant of opening thereof, the change in the outer circumferential side rotor A flow rate occurring subsequent to this drop in flow rate is negligible. Strictly speaking, very small rises occur consequent to increases in the number of revolutions.

Because the opening portion 131 of the second discharge passage 13 of the inner circumferential side rotor B is fully closed the “pressure” of the pump main body (sum of the outer circumferential side rotor A and outer circumferential side rotor B) is equivalent to the pressure of the outer circumferential side rotor A alone. While the change in the pressure of the outer circumferential side rotor A is negligible due to the opening portion 121 of the first return passage 12 being open, strictly speaking, only very small rises in pressure occur consequent to increases in the number of revolutions. In addition, for the “flow rate” of the pump main body, because the opening portion 131 of the second discharge passage 13 of the outer circumferential side rotor B is fully closed, the “flow rate” of the outer circumferential side rotor A constitutes the overall pump flow rate. While the change in the pressure of the outer circumferential side rotor A is negligible due to the opening portion 121 of the first return passage 12 being open, strictly speaking, only very small rises in pressure occurs consequent to increases in the number of revolutions.

While the invention of the subject application constitutes an oil pump pressure control device as described above, it may also constitute a variable flow rate oil pump. This is an oil pump comprising two discharge ports that uses three rotors as means for as means for provision of the two discharge sources thereof. In addition, at times of high revolution when the amount of power consumed by the pump is high, because a discharge port 130 or the second discharge passage 13 of the outer circumferential side rotor B are closed, the outer circumferential side rotor A and the outer circumferential side rotor B are disengaged. Because the flow rate and the pressure of the outer circumferential side rotor B no longer have any effect at all on the flow rate and pressure of the pump main body, even if the flow rate and pressure of the outer circumferential side rotor B are regulated with the aim of increasing efficiency, this has no effect at all on the pump characteristics and, accordingly, allows for the increased degree of design freedom thereof.

Oil pump pressure control device

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CPC

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