US20050196273A1

US20050196273A1 - Power tool

Info

Publication number: US20050196273A1
Application number: US11/068,925
Authority: US
Inventors: Tomomasa Nishikawa; Junichi Toukairin
Original assignee: Hitachi Koki Co Ltd
Current assignee: Koki Holdings Co Ltd
Priority date: 2004-03-04
Filing date: 2005-03-02
Publication date: 2005-09-08
Also published as: CN1663749A; JP2005246542A; DE102005009748A1; CN100348373C

Abstract

Blades are provided on a fan body such that an outer diameter d2 of the fan body is in a range of 45 mm≦d2≦50 mm, an axial height h1 at a substantially intermediate position of the blade is in a range of 0.2≦h1/d2≦0.3 with respect to the outer diameter of the fan body, and an axial height h2 at an outer peripheral edge of the blade is in a range of 0.12≦h2/d2≦0.17 with respect to the outer diameter d2 of the fan body.

Description

BACKGROUND OF THE INVENTION

1. Field of the Invention
The present invention relates to a power tool having a centrifugal fan for cooling a motor.
2. Description of the Related Art
Because of demand for compact size, high output, and low noise of power tools, there has been an increasing need for development of motor cooling fans which are compact and have a high cooling capacity and low noise as well as peripheral components of the fan. Accordingly, optimization of the configuration of blades has been proposed for the purpose of increasing the air volume and reducing the noise (e.g., refer to JP-A-10-153194).
As shown in FIGS. 10 and 11, a plurality of blades 222 of a centrifugal fan 220 disclosed in JP-A-10-153194 are provided in such a manner as to extend from a predetermined radial position to an outer rim on one surface side of the fan body 221, to be arranged at predetermined pitches along a circumferential direction of the fan body 221, and to project in an axial direction of an unillustrated drive shaft from the one surface side.
If the centrifugal fan 220 is rotated, energy is imparted to air by the centrifugal action, and the air passes from an inlet portion at inner ends of the blades 222 through air passages formed by the blades 222 and a fan guide 211, and is exhausted radially outwardly from an outlet portion at outer peripheral portions of the blades 222.
Here, the centrifugal fan 220 is constructed such that the product of a diameter D1 of inner ends of the pair of blades 222 located in the same diametrical direction, a projecting length H1 at the inner end of the blade 222, and an interval L1 in the circumferential direction of the fan body 221 between the mutually opposing portions at the inner ends of the blades 222, and the product of a diameter D2 of the fan body 221, a projecting length H2 at the outer periphery of the blade 222, and an interval L2 in the circumferential direction of the fan body 221 between the mutually opposing portions at the outer edges of the blades 222, become substantially equal. Namely, by ensuring such that D1×H1×L1≅D2×H2×L2, air flows smoothly and the noise reduction effect is exhibited.

SUMMARY OF THE INVENTION

However, the demand for the low noise of power tools is becoming increasingly high, and there is a need to optimize the configuration of the centrifugal fan and make the noise smaller. In general, it is said that the noise of a fluid is proportional to about the sixth power of the flow velocity. In the case of the centrifugal fan, if the number of revolutions is the same, if the outer diameter of the fan is made small, the flow velocity becomes small, so that if the outer diameter of the fan is made small, it is possible to make the noise small. Also, although if the outer diameter of the fan is made small, the flow rate becomes small, but by making the blade height of the fan high, it is possible to compensate for the flow rate by increasing the amount of air to which the centrifugal force is imparted. However, if the blade height is made high, a need arises to form deeper slots in portions of a mold of the fan corresponding to the blades. Since processing becomes difficult such as due to the fact that an end mill is made liable to run out during the formation of the slots, a problem arises in that the manufacturing cost becomes substantially high.
It is an object of the invention to provide a power tool capable of setting an optimum blade height with respect to the fan diameter of the centrifugal fan, and realize low noise and an increase in the air volume at an appropriate manufacturing cost.
According to one aspect of the invention, there is provided with a power tool including: a housing in which an inlet port for introducing air and an outlet port for exhausting the air are formed; a motor having a rotor and a stator, which is accommodated in the housing; and a centrifugal fan capable of rotating with the rotor, which is fixed coaxially to the rotor, the centrifugal fan including; a disc-shaped fan body; and a plurality of blades capable of flowing the air along an axial direction of the rotor radially and outwardly of the fan body, the blades extending from a predetermined position in a radial direction of the fan body to an outer peripheral edge of the fan body, and the blades formed at predetermined pitches along a circumferential direction of the fan body; a first channel formed between the stator and the housing; and a second channel formed between the stator and the rotor, wherein a value S0 is defined by a sectional area which is smallest in the first channel and the second channel among cross sections which are perpendicular to an axial direction of the rotor and are arranged in an axial direction of the rotor, the value S0 is in a range of 350 mm²≦S₀≦650 mm², an outer diameter d2 of the fan body is in a range of 45 mm≦d₂≦50 mm, and a height h1 in the axial direction of the blade at a position where the blade is highest with respect to the outer diameter of the fan body is in a range of 0.2≦h₁/d₂≦0.3.
By thus configuration, the outer diameter d2 of the fan body is in a range of 45 mm≦d₂≦50 mm, and the height h1 in the axial direction of the blade at the position where the blade is highest is in a range of 0.2≦h₁/d₂≦0.3 with respect to the outer diameter of the fan body. Therefore, it is possible to realize low noise and an increase in the air volume at an appropriate manufacturing cost.
According to another aspect of the invention, the height h1 in the axial direction of the blade at the position where the blade is highest is in a range of 0.25≦h₁/d₂≦0.3 with respect to the outer diameter d2 of the fan body.
By thus configuration, the height h1 in the axial direction of the blade at the position where the blade is highest is in a range of 0.25≦h₁/d₂≦0.3 with respect to the outer diameter d2 of the fan body. Therefore, it is possible to realize low noise and an increase in the air volume at a more appropriate manufacturing cost.
According to another aspect of the invention, a height h2 in the axial direction of the blade at an outer peripheral edge of the blade is in a range of 0.12≦h₂/d₂≦0.17 with respect to the outer diameter d2 of the fan body.
By thus configuration, the height h2 in the axial direction of the blade at an outer peripheral edge of the blade is in a range of 0.12≦h₂/d₂≦0.17 with respect to the outer diameter d2 of the fan body. Therefore, it is possible to realize a centrifugal fan which is capable of generating a large air volume and is low-noise.
According another aspect of the invention, a number of the blades, n, is in a range of 23≦n≦30.
By thus configuration, since the number of the blades, n, is in a range of 23≦n≦30, vortices which cause noise are substantially not produced, and it is possible to ensure the passage of air sufficiently. Hence, it is possible to lower the noise while securing a sufficient air volume.
According to another aspect of the invention, the number of the blades, n, is in a range of 25≦n≦28.
By thus configuration, since the number of the blades, n, is in a range of 25≦n≦28, it becomes possible to further lower the noise while securing a more sufficient air volume.
According to another aspect of the invention, a first area S1 is defined by product of both a distance L1 and a height h1, where the distance L1 is defined along a circumferential direction of the fan body and between opposing portions of mutually adjacent blades where the blades are highest, and the height h1 of the blades is defined in an axial direction of the blades where the blades are highest, an inside diameter d1 is defined by a distance between a pair of the blades located in a same diametrical direction of the fan body where the pair of blades are highest, and a second area S2 is defined by a product of a distance L2 and a height h2, where the distance L2 is defined along a circumferential direction of the fan body and between opposing portions of the mutually adjacent blades at outer peripheral edges thereof, and the height h2 is defined in an axial direction of the blade at the outer peripheral edge thereof, d2 is defined by an outer diameter of the fan body, and S1, S2, d1, and d2 are arranged to satisfy a relationship of S₁·d₁=(1±0.3)S₂·d₂.
By thus configuration, since an arrangement is provided so as to satisfy a relationship of S₁·d₁=(1±0.3)S₂·d₂, the air flow between mutually adjacent blades is difficult to be disturbed, making it possible to lower the noise.
According to another aspect of the invention, the blades are configured by an inward portion extending from the predetermined radial position to the position where the blades are highest and an outward portion extending from the position where the blades are highest to the outer peripheral edge, a direction extending toward the outward portion of the blades is inclined by a first predetermined angle α₁in an opposite direction to a rotating direction of the fan with respect to a straight line connecting a center of the fan and an outer peripheral edge of the outward portion of the blades, a direction extending toward the inward portion of the blades is inclined by a second predetermined angle α₂in the opposite direction to the rotating direction of the fan with respect to a straight line connecting the center of the fan and the predetermined radial position, the first predetermined angle α₁is in a range of 30°≦α₁≦50°, and the second predetermined angle α₂is in a range of 0°≦α₂≦10°.
By thus configuration, since the first predetermined angle α1 is in a range of 30°≦α₁≦50°, the velocity of air in the vicinity of the outer peripheral edge of the centrifugal fan can be set to an appropriate velocity, and it is possible to realize low noise while maintaining a sufficient air volume. In addition, since the second predetermined angle α2 is in a range of 0°≦α₂≦10°, it is possible to alleviate a stress occurring at a root of each blade, thereby making it possible to prevent the destruction of the blades. In addition, it is possible to suppress the occurrence of turbulence which causes the noise.
According to another aspect of the invention, the first predetermined angle α₁is in a range of 35°≦α₁≦45°, and the second predetermined angle α₂is in a range of 2.5°≦α₂≦7.5°.
Since the first predetermined angle α1 is in a range of 35°≦α1≦45°, and the second predetermined angle α2 is in a range of 2.5°≦α2≦7.5°, it is possible to realize lower noise while maintaining a more sufficient air volume.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view of a power tool according to one embodiment of the invention which is applied to a grinder;
FIG. 2 is a cross-sectional view taken along line II-II in FIG. 1;
FIG. 3 is a cross-sectional view of a centrifugal fan provided in the power tool;
FIG. 4 is a front elevational view of the centrifugal fan provided in the power tool;
FIG. 5 is a diagram illustrating changes in the flow rate of air in a case where a pressure difference is applied between an inlet and an outlet of air;
FIG. 6 is a diagram illustrating changes in the flow rate of air in a case where a first blade height ratio is varied;
FIG. 7 is a diagram illustrating relationships between, on the one hand, respective combinations of the first blade height and a second blade height and, on the other hand, a noise ratio and an air volume;
FIG. 8 is a diagram illustrating the relationship between the number of blades and an air volume ratio;
FIG. 9 is a cross-sectional view of a modification of the centrifugal fan provided in the power tool;
FIG. 10 is a front elevational view of a conventional centrifugal fan; and
FIG. 11 is a cross-sectional view of the conventional centrifugal fan.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring to FIG. 1, a description will be given of a power tool according to one embodiment of the invention is applied to a grinder (disc grinder).
FIG. 1 shows an overall structure of a disc grinder 1. If it is assumed that the left-hand side in the drawing is a front end, a resin-made handle portion 2, a resin-made motor housing 3, an aluminum alloy-made gear cover 4 are consecutively connected in that order from the rear side, thereby constituting a housing. Spaces defined in the respective interiors of the handle portion 2, the motor housing 3, and the gear housing 4 communicate with each other. A power supply cable 5 is attached to the handle portion 2, and a switch mechanism 6 is incorporated therein. The switch mechanism 6 is provided with a lever 2A which can be operated by a user. The power supply cable 5 connects the switch mechanism 6 to an external power source (not shown), and the connection and disconnection between the switch mechanism 6 and the power source are changed over. In addition, a first air inlet 2 a is formed at a rear end portion of the handle portion 2, and unillustrated second and third air inlets are formed at a front end portion thereof.
A motor 9 having a rotor 7 and a stator 8 is accommodated in the motor housing 3, and the rotor 7 has a drive shaft 10 axially therewith. A fan guide 11 is fixed to the motor housing 3 in front of the motor 9.
Inside the gear housing 4 and in front of the fan guide 11, a centrifugal fan 20 is fixed to the drive shaft 10 concentrically therewith and is provided rotatably with the drive shaft 10. A first air outlet 4 a, a second air outlet 4 b, and an unillustrated third air outlet are formed in the gear cover 4 at positions located radially outwardly of the centrifugal fan 20. In addition, a power transmission mechanism including a pinion gear 12 fixed to one end of the drive shaft 10 and a gear 14 fixed to an output shaft 13, i.e., an output portion, is disposed inside the gear housing 4. The pinion gear 12 meshes with the gear 14 to transmit the rotation of the rotor 7 to the output shaft 13. A grinding wheel 15 is fixed to the output shaft 13.
Referring next to FIG. 2, a description will be given of the internal structure of the motor housing 3. FIG. 2 is a cross-sectional view taken along line II-II of FIG. 1. As described above, the motor 9 has the rotor 7 and the stator 8. The stator 8 is fixedly held in the motor housing 3, and a hollow portion 8 a for loosely inserting the rotor 7 is formed in the stator 8. Further, a plurality of first air channels 3 a are respectively defined by the motor housing 3 and the stator 8, and a plurality of second air channels 3 b are respectively defined by the stator 8 and the rotor 7.
Next, a description will be given of the operation of the disc grinder 1. By pressing the lever 2A against the handle portion 2, an electric current is supplied from the unillustrated external power source to the motor 9 to rotate the rotor 7. The drive shaft 10 also rotates together with the rotation of the rotor 7, and the rotation is transmitted to the output shaft 13 and the grinding wheel 15 by the pinion gear 12 and the gear 14. As the rotating grinding wheel 15 is pressed against a workpiece, the grinding operation is carried out.
At this time, air flows in a built-up pressure space 20 a, which will be described later, as shown by arrows c1, by the rotation of the centrifugal fan 20 fixed to the stator 7, and the pressure drops on the inner diameter side of the centrifugal fan 20, while the pressure becomes high on the outer diameter side thereof. Consequently, air is introduced, as shown by arrows a1, a2, and a3, through the first air inlet 2 a and the unillustrated second and third air inlets in the tail cover 2. Next, the air flows through the first air channel 3 a and the second air channel 3 b, as shown by arrows b1 and b2, to cool the motor 9. Subsequently, the air flows through the built-up pressure space 20 a, as shown by the arrows c1, and flows out to the outside from the first air outlet 4 a, the second air outlet 4 b, and the unillustrated third air outlet, as shown by arrows e1, e2, and e3.
Next, referring to FIG. 3, a description will be given of the structure of the centrifugal fan 20. FIG. 3 is a cross-sectional view of the centrifugal fan 20, and FIG. 4 is a front elevational view, taken from a direction IV in FIG. 3, of the centrifugal fan 20. The centrifugal fan 20 has a fan body 21 and a plurality of blades 22 provided integrally with the fan body 21 and projecting in the axial direction of the fan body 21, and rotates in the direction indicated by arrow A (FIG. 4). The fan body 21 is disc-shaped and is made up of a hub 21A having a rotor fitting/inserting hole 21 a for engaging with the drive shaft 10, as well as a main plate 21B The plurality of blades 22 extend from a predetermined radial position B on the fan body 21 to an outer peripheral edge thereof and are formed at predetermined pitches along the circumferential direction of the fan body 21, in order to allow the air flowing along the axial direction of the rotor 7 to flow radially outwardly of the fan body 21.
As shown in FIG. 4, the blades 22 are inclined in an opposite direction to a rotating direction A with respect to a direction extending radially outwardly from the predetermined radial position B on the fan body 21. Each of the blades 22 is made up of an inward portion 22A extending from the predetermined radial position B to a substantially intermediate position C and an outward portion 22B extending from the substantially intermediate position C to the outer peripheral edge. The inward portion 22A has a configuration in which its axial height becomes gradually higher toward its radially outward side. On the other hand, the outward portion 22B has a configuration in which its axial height becomes gradually lower toward the radially outward side. In addition, the built-up pressure space 20 a where the air flows is defined by the mutually adjacent blades 22, and part of the built-up pressure space 20 a is opposed to the fan guide 11 (FIG. 1).
Here, it is assumed that a distance between the substantially intermediate positions C of the pair of blades 22 located on mutually opposite sides in the same diametrical direction (hereafter referred to as the fan inside diameter) is d1, that the diameter of the fan body 21 (hereafter referred to the fan outer diameter) is d2, that the axial height at the substantially intermediate position C of the blade 22 (hereafter referred to as the blade inward height) is h1, and that the axial height at the peripheral edge of the outward portion 22B of the blade 22 (hereafter referred to as the blade outward height) is h2. Further, it is assumed that an angle formed by a direction in which the outward portion 22B extends and a straight line connecting the center of the centrifugal fan 20 and the outer peripheral edge of the outward portion 22B is α1, and that an angle formed by a direction in which the inward portion 22A extends and a straight line connecting the center of the centrifugal fan 20 and the predetermined radial position B is α2. Furthermore, it is assumed that the distance along the circumferential direction of the centrifugal fan 20 between mutually opposing portions at the substantially intermediate positions C of the mutually adjacent blades 22 (hereafter referred to as the distance between the substantially intermediate positions C) is L1, and that the distance along the circumferential direction of the centrifugal fan 20 between the mutually opposing portions at the peripheral edges of the mutually adjacent blades 22 (hereafter referred to as the distance between the outer peripheral edges of the blades) is L2.
In this embodiment, settings are provided such that the fan inside diameter d1=35 mm; the fan outer diameter d2=48 mm; the blade inward height h1=13 mm; and the blade outward height h2=7 mm. Meanwhile, the dimensions of a conventional centrifugal fan were such that the fan inside diameter d1′=33 mm; the fan outer diameter d2′=52 mm; the blade inward height h1′=9 mm; and the blade outward height h2′=3.5 mm. In addition, settings are provided such that α1=40°, α2=5°, and the number of the blades 22 is set to 27. It should be noted that since there are measurement errors and variations in dimensional values, the symbol ‘=’ when expressing a dimensional value will be used as substantially meaning ‘≅.’
Next, a description will be given of the reason for changing the fan outer diameter d2 from the conventional 52 mm to 48. The selection of the fan outer diameter d2 is based on the property that the relationship between the sound pressure p [pa] of the noise generated by a fluid and the flow velocity v [m/sec] generally becomes
p∝v⁶.
In the centrifugal fan, the relationship between the fan outer diameter d2 and the flow velocity v at the fan outlet generally becomes
d2∝v.
Therefore, if these are combined, we have
p∝d2 ⁶.
In addition, the relationship between the conventional fan outer diameter d2′ and the sound pressure p′ similarly becomes
P′∝d2′⁶.
Accordingly, we have
(p/P′)∝( d 2/ d 2′)⁶.
Namely, the selection of the fan outer diameter d2 is made on the basis of the property that if the fan outer diameter d2 is made small, the flow velocity v becomes small in proportion to it, and the sound pressure p becomes small substantially in proportion to its 6th power. In this embodiment,
(p/P′)∝(48/52)⁶≅0.62.
As such, theoretically, the sound pressure p becomes smaller than the conventional level by approximately 0.62-fold. In addition, as a result of conducting experiments, it was found that in contrast to the fact that a conventional noise value was approximately 81 dB, the noise value in this embodiment was approximately 77.7 dB and became smaller by approximately 3.5 dB. In addition, if the fan outer diameter d2 is in the range of 45 mm≦d2≦50 mm, a substantially similar effect can be obtained.
It should be noted that the sizes of the centrifugal fan 20 in accordance with this embodiment and the conventional centrifugal fan are so designed as to satisfy a formula (1) which will be described later and is for smoothening the flow of air in the centrifugal fan and reducing noise. Accordingly, the fact that the noise value became smaller by approximately 3.5 dB is derived only from the effect of changing the fan outer diameter d2 to 48 mm.
Next, a description will be given of the reason for providing the setting as the blade inward height h1=13 mm with respect to the fan outer diameter d2=48 mm. If we take a look at the ratio between the blade inward height h1 and the fan outer diameter d2 (hereafter referred to as the first blade height ratio), h1/d2, we note that h1/d2=0.27, which is set to be larger than the conventional h1′/d2′=9/52≅0.17. The reason for this is to make the fan outer diameter d2 small so as to compensate for a decline in the flow rate attributable to the fact that the flow velocity became small.
Hereafter, a description will be given of factors affecting the flow rate. The pressure difference P [Pa] between an inlet and an outlet of the built-up pressure space 20 a, which is an air flow capacity necessary for the occurrence of a flow rate Q [m³/min] in the channel from the inlet (the first air inlet 2 a and the unillustrated second and third inlets) in the air channel to the outlet (the first air outlet 4 a, the second air outlet 4 b, and the unillustrated third air outlet), can be expressed by the following formula:
P=aQ²,
where a is a coefficient of channel resistance. In addition, the above formula can be rewritten as
Q={square root}(P/a).
Namely, the factors affecting the flow rate Q are the pressure difference P and the coefficient of channel resistance a. Hereafter, a description will be given of the coefficient of channel resistance a. The coefficient of channel resistance a is a characteristic value which is determined by the configuration of the channel. It is known that the value of the coefficient of channel resistance a is substantially determined by the size of a sectional area (which is set to S0) of the narrowest portion in the channel.
The minimum sectional area S0 in the cooling channel of the disc grinder 1 in this embodiment is a value in which the sectional area is smallest in the first air channel 3 a and the second air channel 3 b among cross sections which are perpendicular to the axial direction of the rotor 7 and are arranged in the axial direction of the rotor 7 (hereafter referred to as the in-motor channel sectional area: S0). The sectional area of the first air channel 3 a is naturally determined from the demand for a compact size for making the outer diameter of the motor housing 3 as small as possible and from the outer diameter of the stator 8 required for obtaining desired power. As for the second air channel 3 b as well, the sectional area is naturally determined from the need to efficiently convert the magnetic force into torque.
As a result of investigating the in-motor channel sectional area S0 with respect to general portable power tools, it was found such that 350 mm²≦S0≦650 mm²or thereabouts. FIG. 5 shows results of investigation into changes in the flow rate Q [mm²] with respect to these in-motor channel sectional areas in cases where the pressure difference between the air inlet and outlet (pressure difference) P [Pa] was produced. The curve X in FIG. 5 was the result in the case of S0=350 mm², and the curve Y was the result in the case of S0=650 mm². In addition, the coefficient of flow resistance a in the curve X was approximately 3000, and the coefficient of flow resistance a in the curve Y was approximately 2000. The coefficient of flow resistance a in the portable power tool of 350 mm²≦S0≦650 mm²is 2000≦a≦3000 or thereabouts.
Hereafter, a description will be given of the air flow capacity of the fan. A factor exerting a large influence on the air flow capacity of the fan is the first blade height ratio. FIG. 6 shows results of investigation into changes in the flow rate Q with respect to the portable power tools of 350 mm²≦S0≦650 mm²in cases where the fan outer diameter was varied in the range of 45 mm≦d2≦50 mm and the first blade height ratio (h1/d2) was varied. In the range of h1/d2≦0.2 in terms of the first blade height ratio, the flow rate increases substantially in proportion to an increase in the first blade height ratio. This is because, in this region, since the channel resistance ratio is sufficiently small with respect to the air flow capacity of the fan, air easily flows through the blades 22, and the flow in the vicinity of the fan is smooth.
In addition, in the region of 0.2≦h1/d2≦0.3, the rate of increase in the flow rate becomes small, and in the range of 0.3≦h1/d2, the flow rate practically ceases to increase. This is because in the range of 0.3≦h1/d2, since the channel resistance ratio with respect to the capacity of the fan is excessively large, air becomes difficult to flow through the blades 22, and the energy of the fan is used to agitate the surrounding air, and because fine bubbles are produced between the blades 22. Namely, in these power tools, if the first blade height ratio h1/d2 is in the range of 0.2≦h1/d2≦0.3, the blade inward height h1 can be set to a proper height in terms of the flow rate and the manufacturing cost.
More preferably, if the first blade height ratio h1/d2 is set in the range of 0.25≦h1/d2≦0.3, the blade inward height h1 can be set to a more appropriate height, and if the first blade height ratio h1/d2 is set such that h1/d2=0.27, the blade inward height h1 can be set to a most appropriate height. It should be noted that in this embodiment, since the fan outer diameter d2 is 48 mm,
h 1/ d 2=0.27.
Hence, we have
h 1=0.27×48≅13 mm,
so that the blade inward height h1 is set as h1=13 mm.
From the above, in a power tool having a cooling channel for cooling a motor by a centrifugal fan and having the in-motor channel sectional area S0 in the range of 350 mm²≦S0≦650 mm², it is possible realize low noise and an increase in the air volume at an appropriate manufacturing cost by setting the fan outer diameter in the range of 45 mm≦d2≦50 mm and the blade inward height h1 in the range of 0.2≦h1/d2≦0.3.
Next, a description will be given of the reason for setting the fan inside diameter d1 to 35 mm and the blade outward height h2 to 7 mm. If it is assumed that an area indicated by the product of the distance between the substantially intermediate positions C, L1, and the blade inward height h1 is an inlet portion area S1 (=L1×h1), and that an area indicated by the product of the distance between the outer peripheral edges of the blades, L2, and the blade outward height h2 is an outlet portion area S2 (=L2×h2), the configuration of the blade 22 is designed so as to satisfy the following relationship:
S 1· d 1=μ· S 2· d 2 0.7≦μ≦1.3 (1)
By so doing, the ratio between the velocity in the radial direction of air between the blades 22 and the velocity in the rotating direction thereof becomes equal between the inlet portion (the substantially intermediate position C of the blade 22) and the outlet portion (the outer peripheral edge of the blade 22), whereby the airflow is made difficult to be disturbed, and the noise can be reduced.
Accordingly, it suffices if the design is made in this embodiment as well so as to satisfy the relationship of Formula (1). Specifically, from Formula (1) (in this embodiment, a setting is provided such that μ=1), we have
(π· d 1/n)· h 1· d 1=(π· d 2/n)· h 2· d 2.
Further, we have
d 1 ² ·h 1= d 2 ² ·h 2 (2)
Here, since in this embodiment d2=48 mm, and h1=13 mm, Formula (2) can be rewritten as
d 1 ²·13=48² ·h 2.
The values of d1=35 mm and h2=7 mm are set so as to satisfy this relationship.
On the basis of the above, the inventors conducted experiments by varying the first blade height ratio h1/d2 and a ratio of the blade outward height h2 to the fan outer diameter d2 (hereafter referred to as the second blade height ratio, h1/d2) in predetermined ranges. FIG. 7 shows results of conducting the experiments by combining the second blade height ratio and the first blade height ratio to (10.0%, 22.0%), (14.5%, 27.0%), and (20.0%, 32.0%), respectively. The ordinate on the left denotes the noise ratio which is a value obtained by dividing the obtained noise value by a predetermined noise value. The ordinate on the right denotes the air volume [m²/min].
From FIG. 7, it can be understood that in the region where the second blade height ratio and the first blade height ratio are less than (12.0%, 25.0%), the noise is small, but it is impossible to obtain a sufficient air volume for cooling the motor 9, and that in the region exceeding (17.0%, 30.0%), a sufficient air volume can be obtained, but the noise becomes large. Accordingly, if the second blade height ratio is in the range of 12.0 to 17.0% and the first blade height ratio is in the range of 25.0 to 30.0%, it is possible to lower the noise while maintaining a sufficient air volume. More preferably, if the second blade height ratio and the first blade height ratio are (14.5%, 27.0%) or in their vicinities, it is possible to realize a centrifugal fan 20 which is capable of producing a large air volume and is low-noise.
Next, a description will be given of the reason for setting the number of the blades 22 to 27. Hereafter, the number of the blades 22 is assumed to be n. As a result of investigating changes in the air volume in cases where the number of the blades 22, n, was changed concerning the centrifugal fan 20 having the fan outer diameter d2 of 45 mm≦d2≦50 mm, the air volume did not change substantially, and a tendency such as the one shown in FIG. 8 was generally shown in the respective cases. The ordinate on the left denotes the air volume ratio which is a value obtained by dividing an air volume value obtained for each number of blades by an air volume value persisting in the case where the number of blades is 27. As can be seen from FIG. 8, the largest air volume was obtained when n=27, and in the range of 23≦n≦30 in terms of the number of blades 22, a very large drop in the air volume was not noted in comparison with n=27. In the range of n<23, since the number of the blades 22, n, is small relative to the fan outer diameter d2, the distance between the adjacent blades 22 becomes large in the vicinity of the outer diameter portion of the centrifugal fan 20. For this reason, the air flow through the blades 22 becomes disturbed, and the air volume declines.
Meanwhile, in the range of n>30, since the number of the blades 22, n, is large relative to the fan outer diameter d2, the interval between the adjacent blades 22 becomes narrow in the vicinity of the inside diameter portion of the centrifugal fan 20. For this reason, air becomes difficult to flow in between the blades 22, so that the air volume declines. For the reasons described above, with the centrifugal fan 20 whose outer diameter d2 is in the range of 45 mm≦d2≦50 mm, if the number of the blades 22, n, is set in the range of 23≦n≦30, it becomes possible to lower the noise while securing a sufficient air volume. More preferably, if the number of the blades 22, n, is set in the range of 25≦n≦28, it becomes possible to further lower the noise while securing a sufficient air volume. Furthermore, if n=27, since it becomes possible to lower the noise most while securing a sufficient air volume, the number of the blades 22 is set to 27 in this embodiment.
Next, a description will be given of the reason for setting α1=40°, α2=5°. The inventors examined changes in the noise and air volume by varying the angles of α1 and α2. As a result, it was found that it is possible to obtain a sufficient air volume and realize low noise in the ranges of 30°≦α1≦50° and 0°≦α2≦10°.
The reason for this is that if α1 is less than 30°, the flow velocity of air becomes fast in the vicinity of the outer peripheral edge of the centrifugal fan 20, which causes noise, whereas if α1 exceeds 50°, the flow velocity of air in the vicinity of the outer peripheral edge of the centrifugal fan 20 becomes slow to the contrary, and a sufficient air volume cannot be obtained. Also, if α2 is less than 0° or exceeds 10°, a large stress easily occurs at the root of the van 22, or turbulence easily occurs, so that it is not desirable. If α2 is in the range of 0°≦α2≦10°, the turbulence can be suppressed.
Furthermore, it was found that in the ranges of 35°≦α1≦40° and 2.5°≦α2≦7.5°, it is possible to obtain a more sufficient air volume and lower noise, and that if α1=40° and α2=5°, it is possible to obtain a largest air volume and reduce the noise to a lowest level.
The power tool is not limited to the above-described embodiment, and various modifications and improvements are possible within the scope of the claims. For example, the centrifugal fan 20 in terms of its configuration may be formed such that, as in a centrifugal fan 120 shown in FIG. 9, a fan body 121 is not disk-shaped, but tapered in such a manner as to be inclined in an opposite direction to the direction in which its blades 122 are projectingly provided. In addition, although the ridge line from the substantially intermediate position C to the outer peripheral edge of the blade 22 has been described as being rectilinear, the invention is not limited to the same, and the ridge line may be formed in the shape of a circular arc or the like as in the centrifugal fan 120. Furthermore, the power tool is not limited to a hammer drill and a disc grinder, and is applicable to a cutter, a screwdriver, and the like.

Claims

1. A power tool comprising:

a housing in which an inlet port for introducing air and an outlet port for exhausting the air are formed;

a motor having a rotor and a stator, which is accommodated in the housing; and

a centrifugal fan capable of rotating with the rotor, which is fixed coaxially to the rotor, the centrifugal fan including;

a disc-shaped fan body; and

a plurality of blades capable of flowing the air along an axial direction of the rotor radially and outwardly of the fan body, the blades extending from a predetermined position in a radial direction of the fan body to an outer peripheral edge of the fan body, and the blades formed at predetermined pitches along a circumferential direction of the fan body;

a first channel formed between the stator and the housing; and

a second channel formed between the stator and the rotor, wherein

a value S0 is defined by a sectional area which is smallest in the first channel and the second channel among cross sections which are perpendicular to an axial direction of the rotor and are arranged in an axial direction of the rotor,

the value S0 is in a range of 350 mm²≦S₀≦650 mm²,

an outer diameter d2 of the fan body is in a range of 45 mm≦d₂≦50 mm , and a height h1 in the axial direction of the blade at a position where the blade is highest with respect to the outer diameter of the fan body is in a range of 0.2≦h₁/d₂≦0.3.

2. The power tool according to claim 1, wherein

the height h1 in the axial direction of the blade at the position where the blade is highest is in a range of 0.25≦h₁/d₂≦0.3 with respect to the outer diameter d2 of the fan body.

3. The power tool according to claim 2, wherein

a height h2 in the axial direction of the blade at an outer peripheral edge of the blade is in a range of 0.12≦h₂/d₂≦0.17 with respect to the outer diameter d2 of the fan body.

4. The power tool according to claim 1, wherein

a number of the blades, n, is in a range of 23≦n≦30.

5. The power tool according to claim 4, wherein the number of the blades, n, is in a range of 25≦n≦28.

6. The power tool according to claim 1, wherein

a first area S1 is defined by product of both a distance L1 and a height h1, where the distance L1 is defined along a circumferential direction of the fan body and between opposing portions of mutually adjacent blades where the blades are highest, and the height h1 of the blades is defined in an axial direction of the blades where the blades are highest,

an inside diameter d1 is defined by a distance between a pair of the blades located in a same diametrical direction of the fan body where the pair of blades are highest, and

a second area S2 is defined by a product of a distance L2 and a height h2, where the distance L2 is defined along a circumferential direction of the fan body and between opposing portions of the mutually adjacent blades at outer peripheral edges thereof, and the height h2 is defined in an axial direction of the blade at the outer peripheral edge thereof,

d2 is defined by an outer diameter of the fan body, and

S1, S2, d1, and d2 are arranged to satisfy a relationship of S₁·d₁=(1±0.3)S₂·d₂.

7. The power tool according to claim 1, wherein

the blades are configured by an inward portion extending from the predetermined radial position to the position where the blades are highest and an outward portion extending from the position where the blades are highest to the outer peripheral edge,

a direction extending toward the outward portion of the blades is inclined by a first predetermined angle α₁in an opposite direction to a rotating direction of the fan with respect to a straight line connecting a center of the fan and an outer peripheral edge of the outward portion of the blades

a direction extending toward the inward portion of the blades is inclined by a second predetermined angle α₂in the opposite direction to the rotating direction of the fan with respect to a straight line connecting the center of the fan and the predetermined radial position,

the first predetermined angle α₁is in a range of 30°≦α₁≦50°, and

the second predetermined angle α₂is in a range of 0°≦α₂≦10°.

8. The power tool according to claim 6, wherein

the first predetermined angle α₁is in a range of 35°≦α₁≦45°, and

the second predetermined angle α₂is in a range of 2.5°≦α₂≦7.5°.