ES2561939T3 - Internally geared oil pump rotor assembly - Google Patents

Abstract

Oil pump rotor assembly comprising: an inner rotor (20) having "n" outer teeth ("n" is a natural number); and an outer rotor (10) having (n + 1) inner teeth that can mesh with the outer teeth, wherein the oil pump rotor assembly is used as an oil pump which, during the rotation of the inner rotors and external, sucks and discharges fluid by changing the volume of cells formed between the internal rotor and the external rotor in which a clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells that has The minimum volume between the cells is designated as "a", a clearance, which is defined between the teeth of the inner and outer rotors that together form one of the cells whose volume increases during the rotation of the inner and outer rotors, it is designates as "b", and a clearance, which is defined between the teeth of the inner and outer rotors that together form one of the cells that has the maximum volume between cells, is designated as "c", characterized in that the following desi Equalities: a <= b <= c, ya <c, and in which when the clearance "b" of the cell located towards the rear as seen in the direction of rotation is also designated as "b1", and the clearance " b "in the cell located forward as seen in the direction of rotation is further designated as" b2 ", the following inequality is satisfied: b1 <= b2 in which a clearance, which is defined between the teeth of the internal rotors and external that together form one of the cells, increases gradually and continuously as the cell is rotationally moved from a position where the volume of the cell is minimized to a position where the volume of the cell is maximized .

Description

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DESCRIPTION

Internally geared oil pump rotor assembly TECHNICAL FIELD

The present invention relates to an oil pump rotor assembly used in an oil pump of the internal gear type that aspirates and discharges fluid by changing the volume of cells formed between an internal rotor and an external rotor.

BACKGROUND OF THE TECHNIQUE

Conventionally, an oil pump of the inner gear type includes an outer rotor having inner teeth, an internal rotor having outer teeth that can be engaged with the inner teeth, and a housing in which an orifice is formed. suction to aspirate fluid and a discharge hole to discharge fluid. The internal rotor is rotated so that the external rotor rotates while the outer teeth engage with the inner teeth, resulting in changes in the volumes of the cells formed between the internal rotor and the external rotor, and thereby aspirated and aspirated. fluid discharge

Each of the cells is independently delimited in a front portion and in a rear portion as seen in the direction of rotation by the outer teeth of the inner rotor and the inner teeth of the external rotor. The volume of each of the cells is minimized in a rotating position in which one of the tooth tips of the outer teeth of the inner rotor coincides positionally with one of the tooth spaces of the inner teeth of the outer rotor and , from this rotation position, the cell aspirates fluid as the volume of the cell increases as it moves over the suction hole. The volume of each of the cells is maximized in a rotational position in which one of the tooth spaces of the outer teeth of the inner rotor coincides positionally with one of the tooth spaces of the inner teeth of the outer rotor, and, from this rotation position, the cell discharges fluid as its volume decreases as it moves over the discharge orifice.

In the oil pump of the inner gear type, the inner rotor is driven so that it rotates, and the outer rotor is rotated because the tooth surfaces of the outer teeth push the tooth surfaces of the inner teeth. In this case, the coupling between the rotors is checked, whereby the rotational force is transmitted. The rotational force is transmitted in the direction substantially perpendicular to the tooth surfaces when the teeth are located near a position in which the cell volume is minimized. On the other hand, when the teeth are located near a position where the cell volume is maximized, because the tooth tips of the rotors come into contact with each other, the rotational force in the direction is not transmitted substantially perpendicular to the tooth surfaces, and the sliding and friction components are predominant.

When the tooth surfaces of the rotors come into contact with each other when the sliding is predominant, the teeth do not contribute to the transmission of the rotational force, and the sliding friction increases due to the contact between the teeth, which can lead to operating noise, and a decrease in mechanical efficiency.

To solve this problem, rotors have been proposed in each of which a recess in the tooth surface is formed to eliminate the contact that does not contribute to the transmission of the rotational force (see, for example, the Japanese patent application not examined, first publication No. Hei 09-166091).

In general, in an oil pump rotor assembly of the inner gear type as mentioned above, clearances are formed between the tooth surfaces of the rotors, which define a cell. The main reason for providing such clearances is to prevent problems in which the rotation of the rotors becomes impossible or noise is emitted because the tooth tips of the rotors interfere with each other due to unwanted shapes and mounting precision of the rotors. , and practical countermeasures have been proposed so that the profiles of the outer rotor teeth are cut evenly, the curve defining the shape of the teeth is partially flattened, or the like.

However, when such clearances are provided merely by taking conventional measures such as uniform cutting of tooth profiles, partial flattening of the tooth surface, or providing recess, unnecessarily increases play between the teeth; Therefore, another problem is that it is difficult to prevent noise due to irregular oscillation of the rotors during rotation.

US 4,976,595 and US 6,077,059 disclose respectively a trocoidal pump and an oil pump.

DISCLOSURE OF THE INVENTION

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The present invention was conceived in view of the above circumstances, and an object of the present invention is to provide an oil pump rotor assembly of the inner gear type that rotates stably without emitting excessive noise.

To achieve the above object, the present invention provides an oil pump rotor assembly that includes: an internal rotor having "n" outer teeth ("n" is a natural number); and an external rotor having (n + 1) inner teeth that can be engaged with the outer teeth, in which the oil pump rotor assembly is used as an oil pump which, during the rotation of the internal and external rotors, aspirates and discharges fluid by changing the volume of cells formed between the internal rotor and the external rotor, in which when a clearance, defined between the teeth of the internal and external rotors that together form one of the cells that has the minimum volume between the cells, is designated as "a", a clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells whose volume increases during the rotation of the internal and external rotors, is designated as "b", and a clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells that has the maximum volume between the cells, is designated as "c", the following desiguations are satisfied altitudes:

a <= b <= c, and a <c

and wherein when the clearance "b" in the cell located backwards as seen in the direction of rotation is further designated as "b1", and the clearance "b" in the cell positioned forward as seen in the direction rotation is also designated as "b2", the following inequality is satisfied:

b1 <= b2

in which a clearance, which is defined between the teeth of the internal and external rotors that form one of the cells together, increases gradually and continuously as the cell moves rotationally from a position where volume is minimized from the cell to a position where the cell volume is maximized.

In a preferred configuration of the previous oil pump rotor assembly, when a clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells whose volume decreases during the rotation of the internal and external rotors, It is designated as "d", the following inequalities are satisfied:

a <= b <= c, a <c ya <= d <= c

and when the clearance "d" in the cell located backwards as seen in the direction of rotation is further designated as "d1", and the clearance "d" in the cell positioned forward as seen in the direction of rotation is also designates as "d2", the following inequality is satisfied:

d1> = d2

In the previous oil pump rotor assembly, the clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells, may gradually decrease as the cell moves rotationally from one position. in which the cell volume is maximized to a position in which the cell volume is minimized.

Therefore, because the clearance between the rotors that form the cell together is minimized in a region of engagement, and then the clearance increases continuously, without decreasing, to a maximum size, the play in a position where minimizes teeth to engage each other, and ensures sufficient clearance in a rotating position where teeth do not contribute to engagement. The outer teeth engage with the inner teeth in a position where a sliding component is minimized so that rotational force is transmitted, and the outer teeth and inner teeth do not contribute to transmitting rotational force in a position in the which increases a slip component. Therefore, an oil pump rotor assembly of the inner gear type can be obtained which does not emit excessive noise while having low friction levels and high mechanical efficiency.

In addition, because in the process in which the cell volume decreases, the clearance between the rotors decreases gradually, without increasing, to a minimum size, sufficient clearance is guaranteed in which the teeth do not contribute to engagement. while the play in which the teeth engage each other is minimized, and therefore an oil pump rotor assembly of the inner gear type that does not emit excessive noise while having low friction levels can be obtained.

In the anterior oil pump rotor assembly, the tooth surfaces of the internal and external rotors can be formed respectively using cycloid curves that are formed by rolling respective rolling circles along respective base circles without sliding.

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In the anterior oil pump rotor assembly, the inner rotor tooth surfaces can be formed using a trocoid envelope curve that is formed by moving a path circle, whose center is located in a trocoid curve, along the trocoid curve. , and the outer rotor tooth tips can be formed using an arc that has the same radius as the path circle.

Accordingly, a rotor assembly of the cycloid type that is formed using cycloid curves and a rotor assembly of the trocoid type that is formed using trocoid curves, both of which have been used in a conventional manner, can be produced so that it emits less noise and that have lower friction levels

In the previous oil pump rotor assembly, each of the internal rotor tooth profiles can be formed such that the tip profile thereof is formed using an epicycloid curve that is formed by rolling a first circumscribed rolling circle Al along a base circle Di without slippage, and the tooth space profile thereof is formed using a hypocycloid curve that is formed by rolling a first inscribed rolling circle B¡ along the base circle DI without slippage, and each one of the outer rotor tooth profiles is formed such that the tip profile thereof is formed using an epicycloid curve that is formed by rolling a second circumscribed rolling circle Ao along a base circle Do without sliding, and the tip profile thereof is formed using a hypocycloid curve that is formed by rolling a second inscribed rolling circle Bo along the base circle Do without sliding, and The internal rotor and the external rotor can be formed so that the following equations are satisfied:

0Bo = 0B¡;

0Do = 0DI (n + 1) / n + t (n + 1) / (n + 2); and 0Ao = 0Ai + t / (n + 2),

where 0D¡ is the diameter of the base circle Di of the internal rotor, 0A¡ is the diameter of the first circumscribed rolling circle Ao, 0B¡ is the diameter of the first inscribed rolling circle B¡, 0Do is the diameter of the base circle Do of the external rotor , 0Ao is the diameter of the second circumscribed rolling circle Ao, 0Bo is the diameter of the second inscribed rolling circle Bo, and t (£ 0) is a gap between the tooth tip of the internal rotor and the tooth tip of the external rotor.

In this case, when the tooth profiles of the Internal and external rotors are determined, because the sum of the bearing distances of the circumscribed rolling circle and the Inscribed rolling circle of the internal rotor must be equal to the circumferential length of the base circle thereof, and the sum of the bearing distances of the circumscribed rolling circle and the inscribed rolling circle of the external rotor must be equal to the circumferential length of the base circle thereof, the following equations must be satisfied:

0D¡ = n (0Ao + 0Bo); Y

0Do = (n + 1) (0Ao + 0Bo);

In addition, in this configuration, the diameters of the Inscribed rolling circles of the internal and external rotors are set to be equal with respect to each other, that is,

0Bo = 0B¡;

to reduce circumferential clearance between the tooth space of the internal rotor and the tooth tip of the external rotor.

The diameter of the base circle of the external rotor is larger than in the case of an oil pump rotor assembly

conventional, that is,

0Do = 0Di (n + 1) / n + n + (n + 1) -t / (n + 2)

Because the total of a multiple of the rolling distance of the circumscribed rolling circle and a multiple of the rolling distance of the inscribed rolling circle must match the circumference length of a base circle, the diameter of the circumscribed rolling circle of the rotor External should be adjusted as follows:

0Ao = 0Ai + t / (n + 2).

According to this oil pump rotor assembly, because an inappropriate radial clearance between the outer teeth of the inner rotor and the inner teeth of the external rotor is guaranteed, and the circumferential clearances between the teeth of the rotors are reduced with respect to which In the conventional case, the vibrations generated between the rotors are reduced, and the silence property of the oil pump can be improved.

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As another configuration of an oil pump rotor assembly, each of the internal rotor tooth profiles can be formed such that the tip profile thereof is formed using an epicycloid curve that is formed by rolling a first rolling circle. circumscribed Di along a base circle "bi" without sliding, and the tooth space profile thereof is formed using a hypocycloid curve that is formed by rolling a first inscribed rolling circle "di" along the base circle " bi "without slippage, and each of the outer rotor tooth profiles is formed such that the tip profile thereof is formed using an epicycloid curve that is formed by rolling a second circumscribed rolling circle Do along a base circle "bo" without sliding, and the tip profile thereof is formed using a hypocycloid curve that is formed by rolling a second rolling circle inscribed "do" as the The base circle “bo” without sliding, and the internal rotor and the external rotor can be formed so that the following equations and inequalities are satisfied:

0b¡ = n (0D¡ + 0d¡);

0bo = (n + 1) (0Do + 0do);

one of 0D¡ + 0d¡ = 2e and 0Do + 0do = 2e;

0Do> 0D¡;

0d¡> 0do; and (0D¡ + 0di) <(0Do + 0do),

where 0b¡ is the diameter of the base circle "bi" of the internal rotor, 0DÍ is the diameter of the first circumscribed rolling circle Di, 0d¡ is the diameter of the first inscribed rolling circle "di", 0bo is the diameter of the base circle "bo "Of the external rotor, 0Do is the diameter of the second circumscribed rolling circle Do, 0do is the diameter of the second inscribed rolling circle" do ", and" e "is the eccentricity distance between the internal and external rotors.

In this case, when the tooth profiles of the internal and external rotors are determined, because the sum of the bearing distances of the circumscribed rolling circle and the inscribed rolling circle of the internal rotor must be equal to the circumferential length of the base circle thereof, and the sum of the bearing distances of the circumscribed rolling circle and the inscribed rolling circle of the external rotor must be equal to the circumferential length of the base circle thereof, the following equations must be satisfied:

0b¡ = n (0D¡ + 0di);

0bo = (n + 1) (0Do + 0do);

The tooth tip profile of the internal rotor that is formed by the first circumscribed rolling circle Di with respect to the tooth space profile of the external rotor that is formed by the second circumscribed rolling circle Do, and the tooth tip profile of the rotor external which is formed by the second inscribed rolling circle "do" with respect to the tooth space profile of the internal rotor which is formed by the first inscribed rolling circle "di" are determined so that the following inequalities are satisfied:

0Do> 0D¡; Y

0d¡> 0do,

so that greater clearance is guaranteed, which is defined between the tooth surfaces of the rotors during engagement. In this case, the play is a separation formed between the tooth surface of the internal rotor, which is opposite to the tooth surface to which force is applied during gearing, and the tooth surface of the rotor

external.

In addition, because the internal rotor and the external rotor engage each other, one of the following equations must be satisfied:

0D¡ + 0d¡ = 2e; Y

0Do + 0do = 2e

In addition, in this invention, in order to make the internal rotor rotate smoothly in the external rotor while ensuring the tip clearance and an appropriate size of the game, and a reduction in gearing resistance, the diameter of the base circle of the external rotor it is made to be larger than that in a conventional case so that the base circle of the internal rotor does not come into contact with the base circle of the external rotor in the engagement region in which the internal rotor engages with the external rotor , that is, the following inequality is satisfied:

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(n + 1) - 0bi <n- 0bo

Therefore, the following inequality is derived:

(0Di + 0di) <(0Do + 0do)

According to the previous configuration, because the circumferential clearances (along the circumference of the base circle) between the tooth surfaces of the rotors are made smaller than in conventional cases while tip clearances are guaranteed between the outer teeth of the internal rotor and the inner teeth of the external rotor, the clearance between the rotors can be reduced, and a silent oil pump can be produced. Specifically, impacts between the inner teeth of the outer rotor and outer teeth of the internal rotor can be prevented even when the driving torque for the oil pump rotor assembly changes while the oil pressure in the oil pump rotor assembly it is low; therefore, the silence property of the oil pump rotor assembly can be guaranteed.

Brief description of the drawings

Figure 1 is a plan view of an oil pump rotor assembly of the inner gear type according to a first embodiment of the present invention, showing the clearances between teeth "a", "b" and "d "

Figure 2 is a plan view of the oil pump rotor assembly of the inner gear type according to the first embodiment of the present invention, in which a gap between teeth "c" is shown.

Figure 3 is a graph in which the clearance between teeth of the oil pump rotor assembly of the inner gear type of the present invention shown in Figure 1 and that of a conventional rotor assembly are compared with respect to the angle Rotation of the internal rotor.

Figure 4 is a plan view showing an oil pump rotor assembly according to a first embodiment of the present invention in which the internal and external rotors thereof satisfy the following

equations:

0Bo = 0B¡;

0Do = 0Di- (n + 1) / n + n + (n + 1) t / (n + 2); and 0Ao = 0Ai + t / (n + 2), and t is set to be 0.12 mm.

Figure 5 is an enlarged view showing the gearing region, indicated by V, of the oil pump shown in Figure 4.

Figure 6 is a graph showing the comparison between the noise from the oil pump incorporating the oil pump rotor assembly shown in Figure 4 and the noise from a conventional oil pump.

Figure 7 is a plan view showing a third embodiment of the oil pump rotor assembly according to the present invention.

Figure 8 is an enlarged view showing the gearing region, indicated by VIII, of the oil pump shown in Figure 7.

Figure 9 is a graph showing a comparison between the set of an oil pump incorporating the oil pump rotor assembly shown in Figure 7 and the set of a conventional oil pump.

Figure 10 is a graph showing a comparison between the noise from an oil pump incorporating the oil pump rotor assembly shown in Figure 7 and the noise from a conventional oil pump.

BETTER WAY TO CARRY OUT THE INVENTION

Next, a first embodiment of the present invention will be explained with reference to Figures 1 to 3.

The oil pump rotor assembly of the inner gear type shown in Figures 1 and 2 is a rotor assembly of the cycloid type in which the teeth of an external rotor 10 and the teeth of an internal rotor 20 are formed using curves respective cycloids, each of which is formed by rolling a rolling circle along

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a base circle The parameters of rotors 10 and 20 are set as follows:

the diameter of the base circle C of the external rotor 10 is 57.31 mm;

the diameter of the circumscribed rolling circle Ao of the external rotor 10 is 2.51 mm;

the diameter of the inscribed rolling circle Bo of the external rotor 10 is 2.70 mm;

the number of teeth Zo of the external rotor 10 is 11 (teeth);

the diameter of the base circle Di of the Internal rotor 20 is 52.00 mm;

the diameter of the circumscribed rolling circle Al of the Internal rotor 20 is 2.50 mm;

the diameter of the inscribed rolling circle B¡ of the internal rotor 20 is 2.76 mm;

the number of teeth Z¡ of the internal rotor 20 is 10 (teeth); Y

the eccentricity distance "e" is 2.60 mm.

The inner rotor 20 is inscribed in the outer rotor 10 while the outer teeth of the inner rotor 20 engage with the inner teeth of the outer rotor 10, so that they form R cells between the teeth. Each of the R cells moves rotationally, while the volume thereof changes when the internal rotor 20 together with the external rotor rotates in the direction indicated by the arrows in Figures 1 and 2 (counterclockwise).

When the rotation position 9 of the Internal rotor 20 is designated as 0o in the lower part of the drawing, and is designated as 180 ° in the upper part of the drawing, the volume of each of the cells R increases gradually, as which rotates the internal rotor 20, from a position in which 9 = 0 ° (figure 1) and its volume (Vmln) is minimized, to a position in which 9 = 198 ° (figure 2) and the volume of it is maximized (Vmax). Each of the R cells aspirates fluid through a suction hole formed in a housing (not shown) during the process in which the volume of the R cell increases.

In this case, a gap between teeth is defined as the region that closes one of the cells R in the circumferential direction, that is, the region in which the separation between the teeth of the rotors 10 and 20 that together form the joint is minimized. cell R.

When a gap between teeth, which is defined between the teeth of the rotors 10 and 20 that together form one of the cells R having the minimum volume (Vmin) between the cells, is designated as "a", a gap between teeth, which is defined between the teeth of the rotors 10 and 20 that together form one of the R cells whose volume increases during the rotation of the rotors 10 and 20, is designated as "b" (Figure 1), and a clearance between teeth , which is defined between the teeth of the rotors 10 and 20 that together form one of the R cells that has the maximum volume (Vmax) between the cells, is designated as "c" (Figure 2), the following inequalities are satisfied:

a <b <c, and a <c

Furthermore, when a gap between teeth, which is defined between the teeth of the rotors 10 and 20 that together form one of the R cells whose volume decreases during the rotation of the rotors 10 and 20, is designated as "d", they are satisfied. The following inequalities:

a <= d <= c.

The comparison between the clearance between the external rotor 10 and the internal rotor 20 in the oil pump rotor assembly of the inner gear type of the present embodiment and that between the rotors in a conventional rotor assembly is shown in the figure 3.

The clearance in the conventional rotor assembly is maximized when the cell volume is minimized, decreases gradually as the cell rotates, and is minimized when the cell volume is maximized. Therefore, in the conventional rotor assembly, the teeth of the rotors tend to come into contact with each other even in beta and gamma zones in which the clearance is smaller than that in an alpha gearing effect zone; therefore, due to friction, mechanical efficiency may decrease, and excessive noise may be emitted.

On the other hand, in the case of the present embodiment, the clearance between teeth between the rotors that together form cell R increases gradually and continuously during the process in which the volume of cell R increases from the minimum volume (Vmin ) up to the maximum volume (Vmax), as shown in Figure 3. More specifically, in relation to the clearance "b" in a range of 0o <9 <198 °, when the clearance "b" in cell R

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located backwards as seen in the direction of rotation is also designated as "b1", and the clearance "b" in the cell R located forward as seen in the direction of rotation is further designated as "b2", is satisfied The following inequality throughout the entire range of rotation position 9:

b1 <= b2.

When the internal rotor 20 rotates from the rotation position 9 = 0 °, the teeth of the external rotor 10 and the teeth of the internal rotor 20 engage each other so that rotational force is transmitted in the area shown in Figure 1. In zone a (i.e. the area of engagement effect), the clearance increases continuously as shown in Figure 3, that is, the clearance in the R cell located forward as seen in the direction of rotation is always greater than the one in cell R located backwards.

The clearance in zone p in which the Internal rotor 20 has been rotated further is greater than that in zone a, and the clearance increases further. Therefore, the teeth of the rotors 10 and 20 tend not to come into contact with each other in the zone P compared to the area of engagement effect a.

The clearance in the zone and (i.e. a performance effect zone) in which the Internal rotor 20 has been rotated adlclonally is greater than that in the zone p, and the clearance increases more, depending on the rotation, up to a maximum value in the position of rotation of the Internal rotor 9 = 198 °. Therefore, the teeth of the rotors 10 and 20 tend not to come into contact with each other in the area and in comparison with the zone p.

The clearance "c" (Figure 2), which is the clearance when the volume of the R cell (Vmax) is maximized, may affect the pump performance because the R cell is at an aspiration transition point a discharge, and the clearance "c" is substantially the same as that in the conventional rotor assembly; therefore, the performance of the pump is not degraded.

In relation to the "d" clearance (figure 1) in the R cell that is forward with respect to the R cell having the maximum volume (Vmax), the "d" clearance decreases gradually, depending on the rotation of the rotor Internal 20, to a minimum value in the rotational position of the Internal rotor 9 = 396 °. In other words, in relation to the "d" clearance in the range of 198 ° <9 <396 °, when the "d" clearance in the backward cell as seen in the direction of rotation is further designated as "d 1 ", and the clearance" d "in the forward cell as seen in the direction of rotation is further designated as" d2 ", the following inequality is satisfied throughout the entire Interval of the rotation position 9:

d1> d2.

Therefore, in the process in which the volume of cell R decreases, as in the process in which the volume of cell R increases, the teeth tend not to come into contact with each other in the area of performance effect and compared to the area of engagement effect a.

As explained above, in the oil pump rotor assembly of the Inner gear type of the present embodiment, the clearance is made to be small in the area of engagement effect at which the rotational force is transmitted from efficiently, the clearance is made to be large in the area of performance effect and in which the rotational force cannot be transmitted efficiently, and the clearance is caused to gradually increase between the areas a and y; therefore, the rotational force is transmitted by contact between the teeth mainly in the area of engagement effect a, and the teeth tend not to come into contact with each other in other areas. As a result, excessive noise and degradation of mechanical efficiency can be prevented.

When the clearance is increased from "a" to "c", it is more preferred that the inequalities a <b, b1 <b2, and b <c be satisfied; however, conditions under which the equations a = b, b1 = b2, or b = c may be partially satisfied provided that an inequality a <c is satisfied, that is, the clearance does not decrease.

Similarly, when the clearance from "c" to "a" is decreased, it is more preferred that the inequalities c> d, d1> d2, and d> a be satisfied; however, conditions under which the equations a = b, b1 = b2, or b = c may be partially satisfied provided that an inequality c> a is satisfied, that is, the clearance does not increase.

In the oil pump rotor assembly of the present embodiment having the dimensions mentioned above, or in the oil pump rotor assembly having dimensions similar to these, it is preferred that the value "a" be in the following Interval :

0.010 mm <a <0.040 mm.

When the “a” value is set to be less than 0.010 mm, the oil pump rotor assembly may not rotate

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gently, and the function as a pump may be lost. On the other hand, when the value "a" is set to be greater than 0.040 mm, the set may become large, and the operating noise may not be reduced.

In addition, it is preferred that the value "c" be in the following range:

0.040 mm <a <0.150 mm.

When the value "c" is set to be less than 0.040 mm, it may become impossible to engage the gearing in the engagement region (at 0 ° in Figure 1). On the other hand, when the value "c is set to be greater than 0,150 mm, oil is excessively leaked through the spacing between the teeth, and the discharge performance of the pump will be extremely degraded.

Next, a second embodiment of the present invention will be explained with reference to Figures 4 to 6.

The oil pump rotor assembly shown in Figure 4 includes an internal rotor 110 provided with "n" outer teeth ("n" indicates a natural number, and n = 10 in this embodiment), and an external rotor 120 provided with " n + 1 ”Inner teeth (n + 1 = 11 in this embodiment) that can be engaged with the outer teeth. The internal rotor 110 and the external rotor 120 are housed in a housing 150.

Between the tooth surfaces of the Internal rotor 110 and the external rotor 120, a plurality of cells C are formed in the direction of rotation of the Internal rotor 110 and the external rotor 120. Each of the cells C is delimited in a front portion and in a rear portion as seen in the direction of rotation of the internal rotor 110 and the external rotor 120 by contact lines between the outer teeth 111 of the inner rotor 110 and the inner teeth 121 of the external rotor 120, and is also delimited in any lateral portion through the housing 150, so that an independent fluid transport chamber is formed. Each of the cells C moves while the Internal rotor 110 and the external rotor 120 rotate, and the volume of each of the cells C increases and decreases cyclically so that a cycle is completed in one rotation.

The Internal rotor 110 is mounted on a rotation axis so that it can rotate around an Oi axis. Each of the tooth profiles of the Internal rotor 110 is formed such that the tooth tip profile thereof is formed using an epicycloid curve that is formed by rolling a first circumscribed rolling circle Ai along a base circle DI of the internal rotor 110 without sliding, and the tooth space profile thereof is formed using a hypocycloid curve that is formed by rolling a first inscribed rolling circle Bi along the base circle DI without sliding.

The external rotor 120 is mounted so that it can rotate, in the housing 150, about an axis Oo that is arranged so that it has a deviation (the eccentricity distance is "e") with respect to the axis Oi. Each of the tooth profiles of the external rotor 120 is formed such that the tooth space profile thereof is formed using an epicycloid curve that is formed by rolling a second circumscribed rolling circle Ao along a base circle C of the external rotor 120 without sliding, and the tooth tip profile thereof is formed using a hypocycloid curve that is formed by rolling a second inscribed rolling circle Bo along the base circle Do without sliding.

When it is assumed that the diameter of the base circle Di of the internal rotor 110, the diameter of the first circumscribed rolling circle Al, the diameter of the first inscribed rolling circle Bi, the diameter of the base circle Do of the external rotor 120, the diameter of the second rolling circle circumscribed Ao and the diameter of the second inscribed rolling circle Bo are 0D¡, 0A¡, 0B¡, 0Do, 0Ao and 0Bo, respectively, the equations to be analyzed below between the internal rotor 110 and the external rotor 120 must be satisfied. Observe that the dimensions will be expressed in millimeters.

First, in relation to the internal rotor 110, because the total of a multiple of the bearing distance of the first circumscribed rolling circle Ai and a multiple of the bearing distance of the first inscribed rolling circle Bi must match the length of the circumference of a base circle, that is, the length of the circumference of the base circle Di of the internal rotor 110 must be equal to the length obtained by multiplying the sum of the rolling distance per revolution of the first circumscribed rolling circle Ai and the distance of bearing of the first inscribed rolling circle Bi by an integer (that is, by the number of teeth of the internal rotor 110),

7- 0DÍ = n-7u (0Ai + 0Bi), that is,

0DÍ = n (0Ai + 0Bi) ■■■ (la).

Similarly, in relation to the external rotor 120, the circumference length of the base circle C of the external rotor 120 must be equal to the length obtained by multiplying the sum of the rolling distance per revolution of the second circumscribed rolling circle Ao and the distance bearing of the second inscribed rolling circle Bo by an integer (that is, by the number of teeth of the external rotor 120),

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Next, the conditions required to determine tooth profiles of the external rotor 120 according to this embodiment will be explained below based on a conventional "ro" external rotor (specifically, the second circumscribed rolling circle "ao" (whose diameter is 0ao), the second rolling circle inscribed "bo" (whose diameter is 0bo) and the base circle "do" (whose diameter is 0do)).

The external rotor "ro" is engaged with the internal rotor 110 according to the present embodiment with a clearance of "t" while it is arranged with respect to the internal rotor 110 so that it has a deviation (the eccentricity distance is "e") . The clearance "t" is a gap formed between one of the tooth tips of the internal rotor 110 and one of the tooth tips of the external rotor 120 in a position that is away from a 180 ° geared region along the direction of rotation when the internal rotor 110 and the external rotor 120 are arranged so that one of the tooth tips of the internal rotor 110 comes into direct contact with one of the tooth spaces of the external rotor 120 in the engagement region.

In this case, the following equations are satisfied:

0do = 0D¡ (n + 1) / n ■■■ (II);

0do = (n + 1) (0ao + 0bo) ■■■ (III);

0ao = 0Ai + t / 2 ■■■ (Illa); Y

0bo = 0Bi-t / 2 ■■■ (IIIb).

The internal rotor 110 that engages with the external rotor "ro" satisfies the following generic equations:

0ai + 0bi = 0AÍ + 0BÍ = 2e ■■ ■ (1); Y

0D¡ = 0do-2e ■■■ (2).

In this embodiment, to reduce circumferential clearances t2 while ensuring radial clearance t1 between the tooth tip of the external rotor 120 and the tooth space of the internal rotor 110 in the engagement region, the diameters are set as follows :

0Bo = 0b¡ = 0B¡ ■■■ (IV).

Based on the previous equations (IV) and (1),

0a¡ = 0A¡ ■■■ (3).

When the inscribed rolling circle of the outer rotor 120 is fixed as described above, the clearance "t" which is expressed as

t = (0Do-0Bo + 0Ao) - (0D¡ + 0Ai + 0Ai) can be expressed, using the equations above (1) to (3) and (IV), as follows:

t = (0Do-0do) - (0Ao-0a¡) ■■■ (V).

Based on the previous equations (Ib), (III), (IV), and (V), t = (0Ao-0ai) (n + 2) (VI); so,

0Ao = 0ai + t / (n + 2).

Next, the diameter 0Do of the base circle Do. Based on the previous equations (Ib) and

0Do-0do = (n + 1) (0Ao + 0Bo) - (n + 1) - (0ao + 0bo).

In addition, based on the previous equations (Illa), (11Ib), and (IV),

0Do-0do = (n + 1) (0Ao-0a¡) ■■■ (Vil).

Using equation (VI), equation (Vil) can be expressed as follows: 0Do-0do = (n + 1) t / (n + 2).

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In addition, using equation (II), 0Do can be expressed as follows:

0Do = (n + 1) - 0Di / n + (n + 1) t / (n + 2) ■■■ (A).

Then, using equation (Ib),

0Ao = 0Do / (n + 1) -0Bo;

therefore, using equation (A),

0Ao = 0Di / n + t / (n + 2) -0Bo,

also, using equations (la) and (IV),

0Ao = 0Ai + t / (n + 2) ■■■ (B).

Summarizing the above equations, the external rotor 120 is formed so that the following equations are satisfied:

0Bo = 0b¡ = 0B¡ ■■■ (IV);

0Do = (n + 1) - 0Di / n + (n + 1) t / (n + 2) ■ (A); and 0Ao = 0Ai + t / (n + 2) ■■■ (B).

Figure 4 shows the oil pump rotor assembly in which the internal rotor 110 is formed so that the above ratio is satisfied (the diameter 0D of the base circle Di is 52.00 mm, the diameter 0AI of the first circumscribed circle Al is 2.50 mm, the diameter 0B¡ of the first inscribed rolling circle Bi is 2.70 mm and the number of teeth Z¡, that is, "n" is 10), the external rotor 120 it is formed so that the previous relationship is satisfied (the external diameter of the same is 70 mm, the diameter 0Do of the base circle Do is 57.31 mm, the diameter 0Ao of the second circumscribed rolling circle Ao is 2.51 mm and the diameter 0Bo of the second inscribed rolling circle Bo is 2.70 mm), and the rotors are combined with the clearance "t" of 0.12 mm, and the eccentricity distance "e" of 2.6 mm.

In the housing 150, a suction hole having a curved shape (not shown) is formed in a region along which each of the cells C moves, which are formed between the rotors 110 and 120, while the volume increases gradually, and a discharge orifice having a curved shape (not shown) is formed in a region along which each of the cells C moves while gradually decreasing the volume of it.

Each of the cells C aspirates fluid as the volume of the same increases when cell C moves through the suction hole after minimizing the volume of cell C in the process of engagement between the outer teeth 111 and the inner teeth 121, and cell C discharges fluid as its volume decreases when cell C moves through the discharge orifice after maximizing the volume of cell C.

Note that if the clearance "t" is too small, a pressure pulse is generated when fluid is discharged from cell C whose volume decreases, which leads to the generation of cavitation noise, thereby increasing the operating noise of the bomb. In addition, the rotors may not rotate smoothly due to the pulsation of pressure.

On the other hand, if the clearance "t" is too large, a pressure pulse is not generated, the operating noise decreases and the sliding resistance between the tooth surfaces due to a large clearance decreases, whereby the mechanical efficiency; however, the fluid tight performance of each of the cells is degraded, and the performance of the pump, specifically, the volumetric efficiency of the pump, degrades. In addition, because the transmission of drive torque in precisely geared positions is not achieved, and the loss in rotation increases, and finally, mechanical efficiency is degraded.

To prevent the above problems, the clearance "t" is preferably adjusted so that the following inequalities are satisfied:

0.03 mm <t <0.30 mm.

In this embodiment, the clearance "t" is set to be 0.12 mm, which is considered to be most preferable.

In the oil pump rotor assembly formed in such a way that the above equations (IV) are satisfied,

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(A), and (B), the profile of the tooth tip of the external rotor 120 and the profile of the tooth space of the internal rotor 110 have substantially the same shape with respect to each other, as shown in Figure 5 As a result, as shown in Figure 5, the circumferential clearances t2 in the engagement phase can be decreased while the radial clearance t1 is guaranteed so that t / 2 is 0.06 mm, which is the same as in conventional rotors; therefore, the gearing impacts between the rotors 110 and 120 during rotation decrease. In addition, because the direction along which the gearing pressure is transmitted perpendicularly to the tooth surfaces, transmission of the torque between the rotors 110 and 120 is performed with high efficiency without slippage, and can reduce heat generation and noise due to slip resistance.

In this embodiment, as in the first embodiment, when a clearance, which is defined between the teeth of the internal and external rotors 110 and 120 that together form one of the cells having the minimum volume between the cells, is designated as "a ", A clearance, which is defined between the teeth of the internal and external rotors 110 and 120 that together form one of the cells whose volume increases during the rotation of the internal and external rotors 110 and 120, is designated as" b ", and a clearance, which is defined between the teeth of the internal and external rotors 110 and 120 that together form one of the cells having the maximum volume between the cells, is designated as "c" (clearances "a", "b" and "c" are not shown), the following inequalities are satisfied:

a <b <c, and a <c.

In addition, when the clearance "b" of the cell located backwards as seen in the direction of rotation is further designated as "b1", and the clearance "b" in the cell positioned forward as seen in the direction of rotation it is also designated as "b2", the following inequality is satisfied:

b1 <b2

Furthermore, when a clearance, which is defined between the teeth of the internal and external rotors 110 and 120 that together form one of the cells whose volume decreases during the rotation of the internal and external rotors 110 and 120, is designated as "d" , the following inequalities are satisfied:

a <b <c, and a <c, and a <d <c.

In addition, when the clearance "d" in the cell located backwards as seen in the direction of rotation is further designated as "d 1", and the clearance "d" in the cell positioned forward as seen in the direction of rotation is also designated as "d2", the following inequality is satisfied:

d1> d2

Figure 6 is a graph showing a comparison between the noise coming from a pump incorporating a conventional oil pump rotor assembly and the noise coming from another pump incorporating the oil pump rotor assembly according to the present embodiment. According to the graph, the noise from the oil pump incorporating the oil pump rotor assembly according to the present embodiment is less than that of the conventional oil pump rotor assembly, that is, the oil pump rotor assembly. Oil of the present embodiment is quieter.

Next, a third embodiment of the present invention will be explained below with reference to Figures 7 to 10.

The oil pump rotor assembly shown in Figure 7 includes an internal rotor 210 provided with "n" outer teeth ("n" indicates a natural number, and n = 10 in this embodiment), and an external rotor 220 provided with " n + 1 ”inner teeth (n + 1 = 11 in this embodiment) that can be engaged with the outer teeth. The internal rotor 210 and the external rotor 220 are housed in a housing 250.

Between the tooth surfaces of the internal rotor 210 and the external rotor 220, a plurality of cells C are formed in the direction of rotation of the internal rotor 210 and the external rotor 220. Each of the cells C is delimited in a front portion and in a rear portion as seen in the direction of rotation of the internal rotor 210 and the external rotor 220 by contact regions between the outer teeth 211 of the inner rotor 210 and the inner teeth 221 of the external rotor 220, and is also delimited in any lateral portion through the housing 250, so that an independent fluid transport chamber is formed. Each of the cells C moves while the internal rotor 210 and the external rotor 220 rotate, and the volume of each of the cells C increases and decreases cyclically so that a cycle in one rotation is completed.

The internal rotor 210 is mounted on a rotation axis so that it can rotate about an axis Oi. Each of the tooth profiles of the inner rotor 210 is formed such that the tooth tip profile thereof is formed using an epicycloid curve that is formed by rolling a first circumscribed rolling circle Di along a base circle " bi "of the internal rotor 210 without sliding, and the tooth space profile thereof is formed

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using a hypocycloid curve that is formed by rolling a first inscribed rolling circle "di" along the base circle "bi" without sliding.

The external rotor 220 is mounted so that it can rotate, in the housing 250, about an axis Oo that is arranged so that it has a deviation (the eccentricity distance is "e") with respect to the axis Oi. Each of the tooth profiles of the external rotor 220 is formed such that the tooth space profile thereof is formed using an epicycloid curve that is formed by rolling a second circumscribed rolling circle Do along a base circle "bo "Of the external rotor 220 without slippage, and the tooth tip profile thereof is formed using a hypocycloid curve that is formed by rolling a second insulated rolling circle" do "along the base circle" bo "without slippage.

When it is assumed that the diameter of the base circle "b" of the internal rotor 210, the diameter of the first circumscribed rolling circle Di, the diameter of the first inscribed rolling circle "di", the diameter of the base circle "bo" of the external rotor 220 , the diameter of the second circumscribed rolling circle Do and the diameter of the second inscribed rolling circle "do" are obi, 0Di, 0di, 0bo, 0Do and 0do, respectively, the equations to be analyzed below between the internal rotor 210 and the external rotor 220. Note that the dimensions in millimeters will be expressed.

First, in relation to the internal rotor 210, because the total of a multiple of the bearing distance of the first circumscribed rolling circle Di and a multiple of the bearing distance of the first inscribed rolling circle "di" must match the length of the circumference of a base circle, that is, the length of the circumference of the base circle "bi" of the internal rotor 210 must be equal to the length obtained by multiplying the sum of the rolling distance per revolution of the first circumscribed rolling circle Di and the bearing distance of the first inscribed rolling circle "di" by an integer (that is, by the number of teeth of the inner rotor 210),

7t- 0b¡ = n-7r (0Di + 0di), that is,

0b¡ = n- (0Di + 0di) ■■■ (la).

Similarly, in relation to the external rotor 220, the length of the circumference of the base circle "bo" of the external rotor 220 must be equal to the length obtained by multiplying the sum of the rolling distance per revolution of the second circumscribed rolling circle Do and the rolling distance of the second rolling circle inscribed "do" by an integer (that is, by the number of teeth of the external rotor 220),

jt 0bo = (n + 1) jr (0Do + 0do), that is,

0bo = (n + 1) (0Do + 0do) ■■■ (Ib).

The tooth tip profile of the internal rotor that is formed by the first circumscribed rolling circle Di with respect to the tooth space profile of the external rotor that is formed by the second circumscribed rolling circle Do, and the tooth tip profile of the rotor external which is formed by the second inscribed rolling circle "do" with respect to the tooth space profile of the internal rotor which is formed by the first inscribed rolling circle "di" are determined so that the following inequalities are satisfied:

0Do> 0D¡; Y

0d¡> 0do,

so that a large clearance that is defined between the tooth surfaces of the rotors during gearing is guaranteed. In this case, the play is a separation formed between the tooth surface of the internal rotor, which is opposite to the tooth surface to which force is applied during gearing, and the tooth surface of the external rotor.

In addition, because the internal rotor and the external rotor engage each other, one of the following equations must be satisfied:

0di + 0di = 2e; Y

0Do + 0do = 2e.

Furthermore, in this invention, in order to make the internal rotor 210 rotate smoothly in the external rotor 220 while ensuring a tip clearance and an appropriate size of the game, and a reduction in the resistance to engagement, the diameter of the base circle "bo" of the external rotor 220 is larger than that in a conventional case so that the base circle "bi" of the internal rotor 210 does not come into contact with the base circle "bo" of the external rotor 220 in the region of gearing in which the internal rotor 210 engages with the external rotor 220, is

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that is, the following inequality is satisfied:

(n + 1) - 0b¡ <n-0bo.

Using this inequality, and equations (la) and (Ib), the following inequality is derived:

(0DI + 0d¡) <(0Do + 0do).

The gearing region mentioned above is a region in which the tooth tip of one of the Inner teeth 221 of the outer rotor 220 is oriented directly towards one of the tooth spaces between the outer teeth 211 of the inner rotor 210.

The internal rotor 210 and the external rotor 220 are formed so that the following inequalities are satisfied:

0.005 mm <(0Do + 0do) - (0Di + 0di) <0.070 mm (hereinafter, (0Do + 0do) - (0Di + 0di) is simply referred to as "A").

In the embodiment, the internal rotor 210 (the diameter 0b¡ of the base circle is 65.00 mm, the diameter 0Di of the first circumscribed rolling circle Di is 3.90 mm, the diameter 0d¡ of the first inscribed rolling circle "di "Is 2.60 mm and the number of teeth" n "is 10) and the external rotor 220 (the external diameter thereof is 87.0 mm, the diameter 0bo of the base circle" bo "is 71,599 mm, the diameter 0Do of the second circumscribed rolling circle Do is 3.9135 mm and the diameter 0do of the second inscribed rolling circle "do" is 2.5955 mm), each of which is formed so that the conditions mentioned above, combine with the eccentricity distance "e" of 3.25 mm to form the oil pump rotor assembly. In this embodiment, the width of the teeth of the rotors (the size in the direction of the axis of rotation) is set to be 10 mm. Because the diameter 0d¡ of the first inscribed rolling circle "di" is set to be 2.60 mm, the diameter 0Do of the second circumscribed rolling circle Do is set to be 3.9135 mm and the diameter 0do of the inscribed second rolling circle "do" is set to be 2.5955 mm, "A" is 0.009 mm (refer to figure 8).

In the housing 250, a suction hole having a curved shape (not shown) is formed in a region along which each of the cells C, which are formed between the rotors 210 and 220, moves while the volume increases gradually, and a discharge orifice that has a curved shape (not shown) is formed in a region along which each of the cells C moves while gradually decreasing the volume of it.

Each of the cells C aspirates fluid as the volume of the same increases when cell C moves through the suction hole after minimizing the volume of cell C in the engagement process between the outer teeth 211 and the inner teeth 221, and cell C discharges fluid as its volume decreases when cell C moves through the discharge orifice after maximizing the volume of cell C.

If "A" is too small, the tip clearance and play cannot be properly fixed, and the engagement noise between the outer teeth 211 of the inner rotor and the inner teeth 221 of the external rotor cannot be reduced.

If "A" is too large, the difference between the height (the size of a tooth along the normal of the base circle) of the outer teeth 211 of the inner rotor and the height of the inner teeth 221 of the outer rotor, and the difference between the width (the size of a tooth along the circumference of the base circle) of the outer teeth 211 of the inner rotor and the width of the inner teeth 221 of the outer rotor cannot be properly fixed; therefore, the game may become zero in some regions during the rotation of internal and external rotors 210 and 220. In this case, the rotors cannot rotate smoothly; therefore, mechanical efficiency can be degraded, and excessive noise can be emitted due to impacts between the outer teeth 211 and the inner teeth 221.

Therefore, it is preferred that "A" be set in the range from 0.005 mm to 0.070 mm, and in this embodiment, "A" is set to be 0.009 mm.

In the configured oil pump rotor assembly as explained above, the tooth tip profile of the external rotor 220 substantially coincides with the tooth space profile of the internal rotor 210. As a result, as shown in the figure 8, the circumferential clearances "ts" along the base circle are made to be small while maintaining the tip clearance "tt" as in a conventional case; therefore, the impacts applied to rotors 210 and 220 during rotation become small. Accordingly, impacts between the inner teeth 221 of the outer rotor 220 and outer teeth 211 of the inner rotor 210 can be prevented even when the driving torque for the oil pump rotor assembly changes while the oil pressure in the assembly of Oil pump rotor is low; therefore, the silence property of the oil pump rotor assembly can be guaranteed. In addition, because the rotational force is transmitted in the direction substantially perpendicular to the tooth surfaces, torque is transmitted between the rotors 210

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and 220 without sliding and with high efficiency, heat and noise generation due to sliding friction can be reduced.

Figure 9 is a graph showing a comparison between a set (shown by a dashed line in Figure 9) of a conventional oil pump rotor assembly with respect to the rotation position of the internal rotor and a set (shown by a continuous line in Figure 9) of the oil pump rotor assembly of the present embodiment with respect to the rotation position of the internal rotor. According to this graph, in the oil pump rotor assembly of the present embodiment, the play in the gearing region, the play in the process in which the volume of cell C increases, and the play in the process in the which decreases the volume of cell C, are smaller than those in the conventional oil pump rotor assembly, and the play in a position in which the volume of cell C is maximized is substantially the same as in The conventional oil pump rotor assembly. Accordingly, in the oil pump rotor assembly of the present embodiment, because the fluid tight performance of cell C having the maximum volume can be guaranteed, and fluid transport efficiency can be maintained substantially equal to the one in a conventional pump. In figure 9, only the game is shown in an internal rotor rotation position from 0 ° to 198 ° because the game in an internal rotor rotation position from 198 ° to 396 ° is similar (symmetrical) to the one from 198 ° to 0o shown in figure 9.

Figure 10 is a graph showing a comparison between the noise coming from an oil pump incorporating a conventional oil pump rotor assembly and the noise coming from the oil pump incorporating the oil pump rotor assembly. The present embodiment. According to this graph, the oil pump rotor assembly of the present embodiment makes it possible to reduce the noise compared to the conventional oil pump rotor assembly, that is, a silent oil pump can be produced, because the play in the gearing region, the play in the process in which the volume of cell C increases, and the play in the process in which the volume of cell C decreases, are smaller than those in the rotor assembly of conventional oil pump as shown in figure 9.

The various elements, dimensions thereof, and combinations thereof explained in the previous embodiments are merely examples, and various modifications can be made according to the design requirements without departing from the scope of the present invention.

For example, in the previous embodiments, the rotors that form the oil pump rotor assembly of the inner gear type are called cycloid rotors having teeth that are formed using cycloid curves; however, any rotor that satisfies the aforementioned slack conditions can be used, such as the so-called trocoid rotors that include an internal rotor having teeth that are formed using a trocoid envelope curve that is formed by moving a circle of trajectory, whose center is located in a trocoid curve, along the trocoid curve, and an external rotor that can be engaged with the internal rotor.

INDUSTRIAL APPLICABILITY

As explained above, according to the oil pump rotor assembly of the inner gear type of the present invention, because the clearance between the rotors that form the cell together is minimized in a region of engagement, and then the clearance continuously increases, without decreasing, to a maximum size, the game in a position in which the teeth are minimized to engage each other, and sufficient clearance is guaranteed in a rotating position in which the teeth do not contribute to the geared

According to another oil pump rotor assembly of the inner gear type of the present invention, because the clearance between the rotors that form together the cell is maximized, and then the clearance decreases continuously, without increasing, to a size minimum in the region of engagement, the play in a position in which the teeth are minimized to engage each other, and sufficient clearance is guaranteed in a rotation position in which the teeth do not contribute to the engagement.

Therefore, the outer teeth engage with the inner teeth in a position where a sliding component is minimized so that rotational force is transmitted, and the outer teeth and inner teeth do not contribute to transmitting rotational force in a position in which a sliding component increases. Therefore, an oil pump rotor assembly of the inner gear type can be obtained which does not emit excessive noise while having low friction levels and high mechanical efficiency.

According to another oil pump rotor assembly of the inner gear type of the present invention, a rotor assembly of the cycloid type that is formed using cycloid curves and a rotor assembly of the trocoid type that is formed using trocoid curves can be produced, which both have been used in a conventional manner, so that they emit less noise and have a lower level of friction; therefore, an oil pump of the inner gear type having high performance can be obtained.

Claims (7)

5 10 fifteen twenty 25 30 35 40 Four. Five fifty 55 60 65 1. Oil pump rotor assembly comprising: an internal rotor (20) having "n" outer teeth ("n" is a natural number); Y an external rotor (10) having (n + 1) inner teeth that can be engaged with the outer teeth, in which the oil pump rotor assembly is used as an oil pump which, during the rotation of the internal and external rotors, aspirates and discharges fluid by changing the volume of cells formed between the internal rotor and the external rotor in which a clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells that has the minimum volume between the cells, is designated as "a", a clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells whose volume increases during the rotation of the internal and external rotors, is designated as "b", and a clearance, which is defined between the teeth of the internal and external rotors that together they form one of the cells that has the maximum volume between the cells, is designated as "c", characterized in that the following inequalities are satisfied: a <b <c, and a <c, and wherein when the clearance "b" of the cell located backwards as seen in the direction of rotation is further designated as "b1", and the clearance "b" in the cell positioned forward as seen in the direction of rotation is also designated as "b2", the following inequality is satisfied: b1 <b2 in which a clearance, which is defined between the teeth of the internal and external rotors that form one of the cells together, increases gradually and continuously as the cell moves rotationally from a position where volume is minimized from the cell to a position where the cell volume is maximized. 2. Oil pump rotor assembly according to claim 1, in which when a clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells whose volume decreases during the rotation of the internal and external rotors, is designated as "d", the following are satisfied inequalities: a <b <c, and a <c, and a <d <c, and wherein when the clearance "d" in the cell located backwards as seen in the direction of rotation is further designated as "d 1", and the clearance "d" in the cell positioned forward as seen in the direction of rotation is also designated as "d2", the following inequality is satisfied: d1> d2 3. Oil pump rotor assembly according to claim 1, wherein the clearance, which is defined between the teeth of the internal and external rotors that together form one of the cells, gradually decreases as it moves rotationally the cell from a position in which the cell volume is maximized to a position in which the cell volume is minimized. 4. Oil pump rotor assembly according to one of claims 1 to 3, wherein the tooth surfaces of the internal and external rotors are formed respectively using cycloid curves that are formed by rolling respective rolling circles along circles respective base without sliding. 5. Oil pump rotor assembly according to one of claims 1 to 3, wherein the inner rotor tooth surfaces are formed using a trocoid envelope curve that is formed by moving a path circle, whose center is located in a trocoid curve, along the trocoid curve, and the outer rotor tooth tips are formed using an arc that has the same radius as the path circle. 6. Oil pump rotor assembly according to claim 1, in which each of the tooth profiles of the internal rotor is formed such that the tip profile thereof is formed using an epicycloid curve that is formed by rolling a first rolling circle. 5 10 fifteen twenty 25 30 35 40 Four. Five fifty 55 circumscribed Ai along a base circle Di without sliding, and the tooth space profile thereof is formed using a hypocycloid curve that is formed by rolling a first inscribed rolling circle Bi along the base circle Di without sliding, and each of the outer rotor tooth profiles is formed such that the tip profile thereof is formed using an epicycloid curve that is formed by rolling a second circumscribed rolling circle Ao along a base circle Do without sliding, and the tip profile thereof is formed using a hypocycloid curve that is formed by rolling a second inscribed rolling circle Bo along the base circle Do without sliding, and in which the internal rotor and the external rotor are formed so that the following equations are satisfied: oBo = 0B¡; 0Do = 0Di (n + 1) / n + t (n + 1) / (n + 2); and 0Ao = 0Ai + t / (n + 2), where 0DÍ is the diameter of the base circle Di of the internal rotor, 0A¡ is the diameter of the first circumscribed rolling circle Ao, 0BÍ is the diameter of the first inscribed rolling circle Bi, 0Do is the diameter of the base circle Do of the external rotor, 0Ao is The diameter of the second circumscribed rolling circle Ao, 0Bo is the diameter of the second inscribed rolling circle Bo, and t (£ 0) is a clearance between the tooth tip of the internal rotor and the tooth tip of the external rotor. 7. Oil pump rotor assembly according to claim 1, in which each of the tooth profiles of the internal rotor is formed such that the tip profile thereof is formed using an epicycloid curve that is formed by rolling a first circumscribed rolling circle Di along a base circle " bi "without sliding, and the tooth space profile thereof is formed using a hypocycloid curve that is formed by rolling a first inscribed rolling circle" di "along the base circle" bi "without sliding, and each of the External rotor tooth profiles are formed such that the tip profile thereof is formed using an epicycloid curve that is formed by rolling a second circumscribed rolling circle Do along a base circle "bo" without sliding, and the The tip profile thereof is formed using a hypocycloid curve that is formed by rolling a second rolling circle Inscribed "do" along the base circle "bo" without sliding, and in which the internal rotor and the external rotor are formed so that the following equations and inequalities are satisfied: 0b¡ = n (0Di + 0di); 0bo = (n + 1) (0Do + 0do); one of 0Di + 0di = 2e and 0Do + 0do = 2e; 0Do> 0D¡; 0d¡> 0do; and (0Di + 0di) <(0Do + 0do), where 0b¡ is the diameter of the base circle "bi" of the Internal rotor, 0Di is the diameter of the first circumscribed rolling circle DI, 0d¡ is the diameter of the first inscribed rolling circle "di", 0bo is the diameter of the base circle "bo "Of the external rotor, 0Do is the diameter of the second circumscribed rolling circle Do, 0do is the diameter of the second inscribed rolling circle" do ", and" e "is the eccentricity distance between the internal and external rotors.