U.S. patent application number 14/360885 was filed with the patent office on 2016-06-16 for hydraulic device.
This patent application is currently assigned to Sumitomo Precision Products Co., Ltd. The applicant listed for this patent is Hiroaki Takeda. Invention is credited to Hiroaki Takeda.
Application Number | 20160169225 14/360885 |
Document ID | / |
Family ID | 50619430 |
Filed Date | 2016-06-16 |
United States Patent
Application |
20160169225 |
Kind Code |
A1 |
Takeda; Hiroaki |
June 16, 2016 |
HYDRAULIC DEVICE
Abstract
A hydraulic device at least has a pair of helical gears 20, 23,
a body 3 containing the gears, bearing members 40, 44 supporting
rotating shafts 21, 24 of the gears 20, 23, and cover plates 7, 8,
11 liquid-tightly fixed to both end surfaces of the body. The cover
plate 8 has a cylinder hole 8a formed at a portion opposite to an
end surface of the rotating shaft 21 of the gear 20 which receives
two thrust forces F.sub.ma, F.sub.pa in the same direction, and a
piston 9 is inserted through the cylinder hole 8a. A working liquid
in a high-pressure side is caused to act on a back surface of the
piston 9 and thereby the piston 9 is pressed onto the end surface
of the rotating shaft 21, thereby causing a drag almost balancing a
resultant force of the two thrust forces F.sub.ma, F.sub.pa to act
on the rotating shaft 21. The thrust forces F.sub.ma, F.sub.pa
acting on the gear 20 are cancelled by this drag.
Inventors: |
Takeda; Hiroaki; (Hyogo,
JP) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Takeda; Hiroaki |
Hyogo |
|
JP |
|
|
Assignee: |
Sumitomo Precision Products Co.,
Ltd
Hyogo
JP
|
Family ID: |
50619430 |
Appl. No.: |
14/360885 |
Filed: |
June 27, 2013 |
PCT Filed: |
June 27, 2013 |
PCT NO: |
PCT/JP2013/067635 |
371 Date: |
May 27, 2014 |
Current U.S.
Class: |
418/191 |
Current CPC
Class: |
F04C 2/084 20130101;
F04C 2/086 20130101; Y10T 74/19953 20150115; F04C 15/0026 20130101;
F04C 2240/52 20130101; F04C 15/0042 20130101; F04C 2/18
20130101 |
International
Class: |
F04C 2/18 20060101
F04C002/18 |
Claims
1. A hydraulic device at least comprising: a pair of helical gears
which each have a rotating shaft provided to extend outward from
both end surfaces thereof, and whose tooth portions mesh with each
other, the pair of helical gears having a tooth profile including
an arc portion at a tooth tip and a tooth root, and forming a
continuous line of contact from one end portion to the other end
portion in a face width direction at a meshing portion; a body open
at both ends and having a hydraulic chamber therein in which the
pair of gears are contained in a state of meshing with each other,
the hydraulic chamber having an arc-shaped inner peripheral surface
with which outer surfaces of the tooth tips of the gears are in
sliding contact; a pair of bearing members which are respectively
disposed on both sides of the gears in the hydraulic chamber of the
body and which support the rotating shafts of the gears so that the
rotating shafts are rotatable; a pair of cover plates which are
respectively liquid-tightly fixed to both end surfaces of the body
to seal the hydraulic chamber, the hydraulic chamber having a
low-pressure side defined on one side of the meshing portion of the
pair of gears and a high-pressure side defined at the other side
thereof; and the body having a flow path which opens into an inner
surface of the low pressure side of the hydraulic chamber and a
flow path which opens into the inner surface of the high pressure
side of the hydraulic chamber, wherein one of the pair of cover
plates which faces a shaft end surface of a thrust-force acting
side of the rotating shaft of one of the gears which receives a
thrust force due to a working liquid in the high-pressure and a
thrust force due to the meshing from the same direction has a
cylinder hole formed at a portion opposite to the shaft end surface
thereof, a flow path for supplying the working liquid in the
high-pressure side into the cylinder hole is formed, a piston is
inserted through the cylinder hole to be capable of being brought
into contact with the shaft end surface opposite to the cylinder
hole, and the working liquid in the high-pressure side is caused to
act on a back surface of the piston to press the piston onto the
shaft end surface, thereby causing a drag approximately balancing a
resultant force of the two thrust forces to act on the shaft end
surface, and on the other hand the one of the pair of cover plates
does not have a cylinder hole formed at a portion opposite to a
shaft end surface of the rotating shaft of the other of the pair of
gears, and the pair of helical gears have a tooth profile
fulfilling a condition that a ratio of contact ratios .epsilon.r
(=.epsilon..beta./.epsilon..alpha.) which is a ratio of overlap
ratio .epsilon..beta. to transverse contact ratio .epsilon..alpha.
is 2<=.epsilon.r<=3.
2. The hydraulic device according to claim 1, wherein the hydraulic
device has seal members with elasticity respectively interposed
between facing surfaces of the pair of cover plates, which face the
pair of bearing members, and facing surfaces of the pair of bearing
members, which face the pair of cover plates, and dividing spaces
between the facing surfaces of the pair of cover plates and the
facing surfaces of the pair of the bearing members, the pair of
bearing members are disposed to be in contact with the end surfaces
of the gears and the working liquid in the high-pressure side is
supplied into the spaces divided by the seal members between the
facing surfaces of the pair of cover plates and the facing surfaces
of the pair of bearing members, and the pair of gears and the pair
of bearing members are configured to be movable in axial directions
of the rotating shafts by elastic deformation of the seal
members.
3. The hydraulic device according to claim 1, wherein the hydraulic
device has a pair of side plates which are respectively interposed
between the pair of gears and the pair of bearing members and which
are disposed to be in contact with the end surfaces of the gears,
and seal members with elasticity respectively interposed between
the pair of side plates and the pair of bearing members to divide
spaces between facing surfaces of the pair side plates, which face
the pair of bearing members, and facing surfaces of the pair of
bearing members, which face the pair of side plates, the working
liquid in the high-pressure side is supplied into the spaces
divided by the seal members between the facing surfaces of the pair
side plates and the facing surfaces of the pair of bearing members,
and the pair of gears and the pair of side plates are configured to
be movable in axial directions of the rotating shafts by elastic
deformation of the seal members.
4. The hydraulic device according to claim 1, wherein the magnitude
of the drag caused to act on the piston is set to be within a range
of 0.9 to 1.1 times of the resultant force of the two thrust
forces.
Description
TECHNICAL FIELD
[0001] The present invention relates to a hydraulic device having a
pair of gears whose tooth surfaces mesh with each other, and
specifically relates to a hydraulic device using, as the pair of
gears, helical gears which have a tooth profile including an arc
portion at a tooth tip and a tooth root, and which form a
continuous line of contact from one end portion to the other end
portion in a face width direction at a meshing portion.
BACKGROUND ART
[0002] Hydraulic devices as mentioned above include a hydraulic
pump which rotates a pair of gears by an appropriate drive motor
and pressurizes a working liquid by the rotational motions of the
gears and discharges the pressurized working liquid, and a
hydraulic motor which rotates gears by introducing a previously
pressurized working liquid therein and uses rotational forces of
rotating shafts of the gears as a power.
[0003] Such a hydraulic device generally has a configuration in
which a pair of gears meshing with each other are contained in a
housing and rotating shafts extended outward from both end surfaces
of each gear are rotatably supported by bearing members which are
contained in the same housing and disposed on both sides of each
gear.
[0004] Conventionally, gears of various shapes have been used as
the pair of gears and some hydraulic devices use helical gears as
the pair of gears. Helical gears have a characteristic that,
because of having a structure in which their teeth are oblique,
gear tooth contact is spread and therefore noise is small, whereas
they have a characteristic that, in a case where they are used as a
hydraulic device, an axial force (thrust force) is generated by
meshing of their teeth and further a thrust force is similarly
generated by the fact that their tooth surfaces receive a pressure
of the working liquid.
[0005] These thrust forces periodically vary due to rotations of
the gears and such periodic variation causes a problem that noise
is generated by vibration of the gears and the bearing members, or
a problem that a gap is formed between the end surfaces of the
gears and the end surfaces of the bearing members by the vibration
and leakage from the high-pressure side to the low-pressure side
through the gap is caused.
[0006] Accordingly, for solving these problems, there has been
suggested a hydraulic device (specifically, a gear pump) configured
to inhibit displacement of the gears in their axial directions by
causing a force in the opposite direction (drag) greater than the
above-described thrust forces to act on the rotating shafts (see
the U.S. Pat. No. 6,887,055 (PTL 1)). A configuration of the gear
pump described in the PTL 1 is shown in FIG. 17.
[0007] As shown in FIG. 17, a gear pump 100 has a body 101 having a
hydraulic chamber 101a formed therein, and a pair of helical gears
115, 120 inserted in the hydraulic chamber 101a with their tooth
portions meshing with each other. As for the pair of gears 115,
120, the gear 115 is a driving gear and the gear 120 is a driven
gear, and their rotating shafts 116, 121 are rotatably supported by
bushes 110a, 110b, 110c and 110d which are similarly inserted in
the hydraulic chamber 101a.
[0008] Further, a front cover 102 is liquid-tightly fixed to the
front end surface of the body 101 by a seal, while an intermediate
plate 106 is similarly liquid-tightly fixed to the rear end surface
of the body 101 by a seal and a rear cover 104 is similarly
liquid-tightly fixed to the rear end surface of the intermediate
plate 106 by a seal. The body 101, the front cover 102, the
intermediate plate 106 and the rear cover 104 together form a
housing within which the hydraulic chamber 101a is sealed. It is
noted that the rotating shaft 116, which is inserted through a
through hole 102a of the front cover 102 and extended outward, is
sealed by a not-shown seal between the outer peripheral surface of
the rotating shaft 116 and the inner peripheral surface of the
through hole 102a.
[0009] The hydraulic chamber 101a is divided in two, a
high-pressure side and a low-pressure side, at a meshing portion of
the pair of gears 115, 120, and when the driving gear 115 is driven
and rotated by an appropriate driving source and the pair of gears
115, 120 thereby rotate, a working liquid is introduced into the
low pressure side through a not-shown intake port and the
introduced working liquid is led to the high pressure side while
being pressurized by an action of the pair of gears 115, 120, and
the high-pressure working liquid is discharged through a not-shown
discharge port.
[0010] Further, the intermediate plate 106 has through holes 106a,
106b bored therethrough at portions corresponding to the rotating
shafts 116, 121, respectively, and pistons 108, 109 are inserted
through the through holes 106a, 106b, respectively. Further, a
concave hydraulic chamber 104a corresponding to a region including
the through holes 106a, 106b is formed in the surface being in
contact with the intermediate plate 106 (front surface) of the rear
cover 104, and the working liquid in the high-pressure side is to
be supplied into the hydraulic chamber 104a through an appropriate
flow path. Furthermore, the working liquid in the high-pressure
side is to be supplied into between the front surface of the
intermediate plate 106 and the rear surfaces of the bushes 110a,
110c through an appropriate flow path.
[0011] According to the gear pump 100 having the above-described
configuration, during the operation of the gear pump 100, the
working liquid in the high-pressure side is supplied into the
hydraulic chamber 104a of the rear cover 104, the pistons 108, 109
are pressed forward by the high-pressure working liquid, and the
gears 115, 120 are pressed forward by the pistons 108, 109 via the
rotating shafts 116, 121, and simultaneously the bushes 110a, 110c
are pressed forward by the high-pressure working liquid supplied
into between the front surface of the intermediate plate 106 and
the rear surfaces of the bushes 110a, 110c. Due to these actions,
the bushes 110a, 110c, the gears 115, 120 and the bushes 110b, 110d
are integrally pressed forward and the bushes 110b, 110d are
pressed onto the rear end surface of the front cover 102.
[0012] It is noted that the pressing force for integrally pressing
a structure comprising the bushes 110a, 110b, the gears 115, 120
and the bushes 110b, 110d forward is set to be greater than the
thrust forces generated by the rotations of the gears 115, 120.
Further, the pistons 108, 109 have their respective pressure
receiving areas (cross-sectional areas) which are respectively
determined in accordance with the thrust forces acting on the
driving gear 115 and the driven gear 120, and the cross-sectional
area of the piston 108 is larger than that of the piston 109.
[0013] As described above, in a hydraulic device using helical
gears, the thrust forces generated by rotations of the helical
gears causes vibration and noise and causes leakage from the high
pressure side to the low pressure side. However, according to the
gear pump 100, since the structure comprising the bushes 110a,
110c, the gears 115, 120 and the bushes 110b, 110d is pressed onto
the rear end surface of the front cover 102 by integrally pressing
them forward with a force greater than the thrust forces, the gears
115, 120 and the bushes 110a, 110b, 110c, 110d do not vibrate and
the occurrence of the above-described noise and leakage problems
caused by vibration is prevented.
[0014] It is noted that as a gear pump using helical gears, besides
the gear pump as disclosed in the PTL 1, conventionally, there have
been known a gear pump as disclosed in the Japanese Unexamined
Patent Application Publication No. H2-95789 (PTL 2) and a gear pump
as disclosed in the Japanese Examined Utility Model Application
Publication No. S47-16424 (PTL 3).
[0015] In the gear pump disclosed in the PTL 2, the pressure of the
fluid to be driven is caused to act on the shaft end surface
opposite the output side of the driving gear to cause a thrust
force acting on the driving shaft due to this pressure and the
thrust force acting on the driving shaft due to meshing of the
gears to cancel each other out.
[0016] Further, in the gear pump disclosed in the PTL 3, similarly
to the gear pump disclosed in the PTL 1, a thrust force due to a
pressure fluid is caused to act on each of the shaft ends of the
driving gear and the driven gear to cause these thrust forces and
the thrust forces acting on the driving gear and the driven gear to
cancel each other out.
CITATION LIST
Patent Literature
[0017] PTL 1: U.S. Pat. No. 6,887,055
[0018] PTL 2: Japanese Unexamined Patent Application Publication
No. H2-95789
[0019] PTL 3: Japanese Examined Utility Model Application
Publication No. S47-16424
SUMMARY OF INVENTION
Technical Problem
[0020] However, the above-described conventional gear pumps have a
problem as described below. That is, first, in the gear pump 100
described in the PTL 1, although the noise and leakage problems
caused by vibration are prevented, there is a problem that, because
the gear pump 100 is configured to always integrally press the
structure comprising the bushes 110a, 110c, the gears 115, 120 and
the bushes 110b, 110d forward with a force greater than the thrust
forces and thereby press it onto the rear end surface of the front
cover 102, the end surfaces of the bushes 110a, 110b, 110c and 110d
are always in sliding contact with the end surfaces of the gears
115, 120 with a considerable pressure, and thereby burn occurs on
the end surfaces of the bushes 110a, 110b, 110c, 110d. Further, if
such a state continues for a long time, finally the end surfaces of
the bushes 110a, 110b, 110c, 110d are damaged and this results in
the occurrence of noise and leakage from the damaged portions, and
further, the worst situation that the gears 115, 120, the bushes
110a, 110b, 110c, 110d, the body 101 and the like are broken can
occur.
[0021] Further, although the gear pump disclosed in the PTL 2 is
configured to cause a hydraulic pressure to act on only a shaft end
of the driving shaft and thereby apply a thrust force corresponding
to the hydraulic pressure to the driving shaft, this thrust force
opposes the thrust force generated by meshing of the driving gear
and the driven gear, and, in this gear pump, the thrust force
generated by hydraulic pressures acting on the driving gear and the
driven gear are not taken into consideration at all. Therefore, in
this gear pump, a periodically varying thrust force cannot be
reduced and it is not possible to appropriately maintain a contact
pressure between the end surfaces of the helical gears and the
members in contact therewith. Therefore, the problem of the
occurrence of noise and leakage is not solved. Further, the PTL 2
only discloses that a thrust force as drag is caused to act on the
driving shaft, and therefore the specific magnitude of drag that
should be caused to act on the driving shaft is not clear at
all.
[0022] On the other hand, the PTL 3 discloses the specific
magnitudes of the two thrust forces acting on the helical gears,
that is, the thrust force generated by meshing and the thrust force
generated by a hydraulic pressure. However, according to knowledge
obtained as a result of eager studies by the inventors, it was
found out that, in a case of using helical gears which have a tooth
profile including an arc portion at a tooth tip and a tooth root
and forming a continuous line of contact from one end portion to
the other end portion in a face width direction at a meshing
portion, the thrust forces acting on them have magnitudes different
from those disclosed in the PTL 3. Therefore, in a case of using
helical gears having such a tooth profile, even if thrust forces as
disclosed in the PTL 3 are caused to act on the rotating shafts, a
periodically varying thrust force cannot be reduced and it is not
possible to appropriately maintain a contact pressure between the
end surfaces of the helical gears and the members in contact
therewith, and therefore the problem of the occurrence of noise and
leakage cannot be solved.
[0023] Further, in the gear pumps disclosed in the PTLs 1 to 3,
mechanical efficiency is not taken into consideration at all, and,
in the case where mechanical efficiency is not taken into
consideration, it is not possible to exactly cancel the thrust
forces acting on the helical gears and the above-described problems
are not completely solved.
[0024] Furthermore, the inventors, as a result of their eager
studies, obtained knowledge that, in the case of using the
above-described helical gears, that is, helical gears which have a
tooth profile including an arc portion at a tooth tip and a tooth
root and forming a continuous line of contact from one end portion
to the other end portion in a face width direction at a meshing
portion, there can be a case where no thrust force acts on the
driven gear.
[0025] The present invention has been achieved in view of the
above-described circumstances and an object thereof is to provide a
hydraulic device using helical gears which have a tooth profile
including an arc portion at a tooth tip and a tooth root and
forming a continuous line of contact from one end portion to the
other end portion in a face width direction at a meshing portion
and which is capable of reducing a periodically varying thrust
force, appropriately maintaining a contact pressure between end
surfaces of the helical gears and members in contact therewith and
preferably maintaining tight contact between them, and effectively
suppressing the occurrence of noise and leakage.
Solution to Problem
[0026] The present invention, for solving the above-described
problem, relates to a hydraulic device comprising:
[0027] a pair of helical gears which each have a rotating shaft
provided to extend outward from both end surfaces thereof, and
whose tooth portions mesh with each other, the pair of gears having
a tooth profile including an arc portion at a tooth tip and a tooth
root, and forming a continuous line of contact from one end portion
to the other end portion in a face width direction at a meshing
portion;
[0028] a body open at both ends and having a hydraulic chamber
therein in which the pair of gears are contained in a state of
meshing with each other, the hydraulic chamber having an arc-shaped
inner peripheral surface with which outer surfaces of the tooth
tips of the gears are in sliding contact;
[0029] a pair of bearing members which are respectively disposed on
both sides of the gears in the hydraulic chamber of the body and
which support the rotating shafts of the gears so that the rotating
shafts are rotatable; and
[0030] a pair of cover plates which are respectively liquid-tightly
fixed to both end surfaces of the body to seal the hydraulic
chamber, wherein
[0031] the hydraulic chamber has a low-pressure side defined at one
side of the meshing portion of the pair of gears and a
high-pressure side defined at the other side thereof, and the body
has a flow path which opens into the inner surface of the
low-pressure side of the hydraulic chamber and a flow path which
opens into the inner surface of the high-pressure side of the
hydraulic chamber.
[0032] Further, the hydraulic device of the present invention has
seal members with elasticity respectively interposed between facing
surfaces of the pair of cover plates, which face the pair of
bearing members, and facing surfaces of the pair of bearing
members, which face the pair of cover plates, and dividing spaces
between the facing surfaces of the pair of cover plates and the
facing surfaces of the pair of the bearing members, and
[0033] the hydraulic device is configured so that: the pair of
bearing members are disposed to be in contact with the end surfaces
of the gears; a working liquid in the high-pressure side is
supplied into the spaces divided by the seal members between the
facing surfaces of the pair of cover plates and the facing surfaces
of the pair of bearing members; and the pair of gears and the pair
of bearing members can be moved in axial directions of the rotating
shafts by elastic deformation of the seal members.
[0034] Alternatively, the hydraulic device of the present invention
has a pair of side plates which are respectively interposed between
the pair of gears and the pair of bearing members and which are
respectively disposed to be in contact with the end surfaces of the
gears, and has seal members with elasticity respectively interposed
between the pair of side plates and the pair of bearing members to
divide spaces between facing surfaces of the pair side plates,
which face the pair of bearing members, and facing surfaces of the
pair of bearing members, which face the pair of side plates, and
further, the hydraulic device is configured so that a working
liquid in the high-pressure side is supplied into the spaces
divided by the seal members between the facing surfaces of the pair
side plates and the facing surfaces of the pair of bearing members
and the pair of gears and the pair of side plates can be moved in
axial directions of the rotating shafts by elastic deformation of
the seal members.
[0035] Further, in the present invention, each of the
above-described hydraulic devices has a configuration in which: one
of the pair of cover plates which faces a shaft end surface of a
thrust-force acting side of the rotating shaft of one of the gears
which receives a thrust force due to the working liquid in the
high-pressure and a thrust force due to the meshing from the same
direction has a cylinder hole formed at a portion opposite to the
shaft end surface thereof; a flow path for supplying the working
liquid in the high-pressure side into the cylinder hole is formed;
a piston is inserted through the cylinder hole to be capable of
being brought into contact with the shaft end surface opposite to
the cylinder hole; and the working liquid in the high-pressure side
is caused to act on a back surface of the piston to press the
piston onto the shaft end surface, thereby causing a drag
approximately balancing a resultant force of the two thrust forces
to act on the shaft end surface, whereas the one of the pair of
cover plates does not have a cylinder hole formed at a portion
opposite to a shaft end surface of the rotating shaft of the other
of the pair of gears thereof.
[0036] As described above, in a hydraulic device using helical
gears, a thrust force is generated due to meshing of the teeth
(hereinafter, referred to as a "meshing thrust force"), and a
thrust force is similarly generated due to the fact that the tooth
surfaces receive a pressure of a working liquid (hereinafter,
referred to as a "pressure receiving thrust force").
[0037] Of these thrust forces, the pressure receiving thrust force
acts on the tooth surfaces of the pair of gears in the same manner,
and therefore the directions of the pressure receiving thrust
forces acting on the pair of gears are the same direction. On the
other hand, since the meshing thrust force is generated due to
meshing of the tooth portions of the pair of gears and the meshing
thrust forces acting on the gears act as a reaction force to each
other, the directions of the meshing thrust forces acting on the
pair of gears are opposite directions. Therefore, the directions of
the meshing thrust force and the pressure receiving thrust force
acting on one gear of the pair of gears are the same and a thrust
force as a resultant force of the meshing thrust force and the
pressure receiving thrust force acts on the one gear. On the other
hand, the directions of the meshing thrust force and the pressure
receiving thrust force acting on the other gear of the pair of
gears are opposite to each other, and a thrust force as a
differential between the meshing thrust force and the pressure
receiving thrust force acts on the other gear.
[0038] Further, according to knowledge of the inventors, in a case
where each of the helical gears is a gear which has a tooth profile
including an arc portion at a tooth tip and a tooth root and
forming a continuous line of contact from one end portion to the
other end portion in a face width direction at a meshing portion
(hereinafter, such a helical gear is referred to as a
"continuous-line-of-contact meshing gear"), and the tooth profile
fulfills the condition that a ratio of contact ratios
.epsilon..sub.r (=.epsilon..sub..beta./.epsilon..sub..alpha.) which
is the ratio of the overlap ratio .epsilon..sub..beta. to the
transverse contact ratio .epsilon..sub..alpha. of the gears is
2<=.epsilon..sub.r<=3, there is a case where the meshing
thrust force and the pressure receiving thrust force have the same
magnitude, and it is possible to achieve a hydraulic device within
a practical mechanical efficiency.
[0039] Thus, in the case where the meshing thrust force and the
pressure receiving thrust force have the same magnitude, the
pressure receiving thrust force and the meshing thrust force are
cancelled out on the other gear and no thrust force acts
thereon.
[0040] On the other hand, in the present invention, since, as
described above, the piston is pressed onto the shaft end surface
of the rotating shaft of the gear on which a resultant force of the
meshing thrust force and the pressure receiving thrust force acts
and thereby a drag having a magnitude which approximately balances
the resultant force is caused to act on the shaft end surface of
the rotating shaft by the piston, no thrust force acts also on the
one gear.
[0041] Thus, in the hydraulic device of the present invention, it
is possible to achieve a state where both of the pair of gears do
not receive a thrust-directional force. Therefore, according to the
present invention, there is not caused the above-described
conventional problem that seizure or damage caused by a thrust
force occurs on the bearing members or the side plates which are
into sliding contact with the end surfaces of the pair of
gears.
[0042] Further, in the hydraulic device of the present invention,
since providing the piston for causing a reaction force to act on
only the rotating shaft of one of the gears achieves the state
where no thrust force acts on both of the gears, the
above-described problem can be solved while reducing costs for
manufacturing the hydraulic device.
[0043] Further, in a case where mechanical efficiency is not taken
into consideration, it is preferred that the
"continuous-line-of-contact meshing gear" has a tooth profile which
fulfills the condition that the ratio of contact ratios
.epsilon..sub.r is 2 or 3. According to knowledge of the inventors,
in a case where it is assumed that an input value and an output
value in the hydraulic device of the present invention are equal to
each other, that is, mechanical efficiency is 100%, when the gears
have a tooth profile which fulfills the condition that the ratio of
contact ratios .epsilon..sub.r is 2 or 3, the hydraulic device is a
hydraulic device having practical gears and it is possible to cause
the meshing thrust force and the pressure receiving thrust force to
have the same magnitude and therefore the above-described effect is
obtained.
[0044] Further, in the present invention, since the working liquid
in the high-pressure side is caused to act on the back surfaces of
the bearing members or the side plates, which are into contact with
both end surfaces of the pair of gears, to bring the bearing
members or the side plates into tight contact with both end
surfaces of the pair of gears, and the pair of gears and the
bearing members or side plates which are brought into tight contact
therewith are provided so that they can be moved in the axial
directions of the rotating shafts by elastic deformation of the
seal members, even if periodic variation occurs on the thrust
forces or sudden vibration occurs on the hydraulic device, such
variation and sudden vibration are absorbed by movement of the pair
of gears and the bearing members or the side plates in the axial
directions of the rotating shafts, and the occurrence of noise
caused by such variation and vibration is suppressed. Further,
since the bearing members or the side plates are brought into tight
contact with both end surfaces of the gears by the working liquid
in the high-pressure side which acts on the back surfaces thereof,
leakage of the working liquid through the end surfaces of the gears
is appropriately suppressed.
[0045] Further, it is preferred that the magnitude of the drag
caused to act on the piston is within a range of 0.9 to 1.1 times
of the resultant force, and this drag is determined in accordance
with a pressure receiving area S (mm.sup.2) of the piston and the
pressure receiving area S (mm.sup.2) of the piston is set so that a
drag within the above-mentioned range is generated.
[0046] It is noted that the "continuous-line-of-contact meshing
gear" in the present invention includes an involute gear, a
sine-curve gear, a segmental gear, a parabolic gear, etc.
Advantageous Effects of Invention
[0047] As described above, according to the present invention, in a
hydraulic device using "continuous-line-of-contact meshing gears"
as gears, the thrust forces acting on the gears can be reduced and
the gears can be brought into a natural state. Therefore, according
to the present invention, there is not caused the above-described
conventional problem that seizure or damage caused by the thrust
forces occurs on the bearing members or side plates being in
sliding contact with both end surfaces of the pair of gears.
[0048] Further, even if periodic variation occurs on the thrust
forces or sudden vibration occurs on the hydraulic device, such
variation and sudden vibration can be absorbed by movement of the
pair of gears and the bearing members or the side plates in the
axial directions of the rotating shafts, and the occurrence of
noise caused to such variation and vibration can be suppressed.
Furthermore, since the bearing members or the side plates are
brought into tight contact with both end surfaces of the gears by
the working liquid in the high pressure side which acts on the back
surfaces thereof, leakage of the working liquid through the end
surfaces of the gears can be appropriately suppressed.
BRIEF DESCRIPTION OF DRAWINGS
[0049] FIG. 1 is a plan sectional view of an oil hydraulic pump
according to one embodiment of the present invention;
[0050] FIG. 2 is a front sectional view taken along the arrows A-A
in FIG. 1;
[0051] FIG. 3 is a plan view of a bush of the oil hydraulic pump
according to the embodiment;
[0052] FIG. 4 is a side view as seen in the direction indicated by
the arrow B in FIG. 3;
[0053] FIG. 5 is an illustration for explaining a meshing thrust
force;
[0054] FIG. 6 is an illustration for explaining a pressure
receiving thrust force;
[0055] FIG. 7 is an illustration for explaining the pressure
receiving thrust force;
[0056] FIG. 8 is an illustration showing a specific mode of meshing
of gears;
[0057] FIG. 9 is an illustration showing a specific mode of meshing
of gears;
[0058] FIG. 10 is an illustration showing a specific mode of
meshing of gears;
[0059] FIG. 11 is an illustration showing a specific mode of
meshing of gears;
[0060] FIG. 12 is an illustration for explaining a pressure
receiving area of a gear;
[0061] FIG. 13 is an illustration for explaining the pressure
receiving area of a gear;
[0062] FIG. 14 is a plan sectional view of an oil hydraulic pump
according to another embodiment of the present invention;
[0063] FIG. 15 is a side view of a bush according to the embodiment
shown in FIG. 14;
[0064] FIG. 16 is a plan sectional view of an oil hydraulic pump
according to a further another embodiment of the present invention;
and
[0065] FIG. 17 is a plan sectional view of a conventional gear
pump.
DESCRIPTION OF EMBODIMENTS
[0066] Hereinafter, a specific embodiment of the present invention
will be described on the basis of the drawings. It is noted that
the hydraulic device of this embodiment is an oil hydraulic pump
and a hydraulic oil is used as working liquid.
[0067] As shown in FIGS. 1 and 2, an oil hydraulic pump 1 has a
housing 2 having a hydraulic chamber 4 formed therein, a pair of
helical gears which are disposed in the hydraulic chamber 4 and
have a tooth profile including an arc portion at a tooth tip and a
tooth root and forming a continuous line of contact from one end
portion to the other end portion in a face width direction at a
meshing portion, that is, a pair of "continuous-line-of-contact
meshing gears" as described above (hereinafter, simply referred to
as gears) 20, 23, bushes 40, 44 as a pair of bearing members, and a
pair of side plates 30, 32.
[0068] The housing 2 comprises a body 3 in which the hydraulic
chamber 4 having a space with a substantially 8-shaped
cross-section is formed from one end surface to the other end
surface thereof, a front cover 7 which is liquid-tightly fixed to
the one end surface (front end surface) of the body 3 via a seal
12, an intermediate cover 8 which is similarly liquid-tightly fixed
to the other end surface (rear end surface) of the body 3 via a
seal 13, and an end cover 11 which is liquid-tightly fixed to a
rear end surface of the intermediate cover 8 via a seal 14, and the
hydraulic chamber 4 is closed by the front cover 7 and the
intermediate cover 8.
[0069] One of the pair of gears 20, 23 is a driving gear 20 and the
other is a driven gear 23, and the driving gear 20 has a
right-handed helical tooth portion and the driven gear 23 has a
left-handed helical tooth portion. The gears 20, 23 respectively
have rotating shafts 21, 24 which are respectively provided to
extend in the axial directions of the gears 20, 23 from both end
surfaces of the gears 20, 23. Further, the pair of gears 20, 23 are
inserted in the hydraulic chamber 4 in a state of meshing with each
other so that outer surfaces of their tooth tips are in sliding
contact with an inner peripheral surface 3a of the hydraulic
chamber 4, and the hydraulic chamber 4 is divided in two, a
high-pressure side and a low-pressure side, at the meshing portion
of the pair of gears 20, 23. Further, an end portion of the
rotating shaft 21 on the front side of the driving gear 20 is
formed in a tapered shape and a screw portion 22 is formed on the
tip thereof, and the end portion of the rotating shaft 21 extends
outward through a through hole 7a formed in the front cover 7 and
an oil seal 10 provides sealing between the outer peripheral
surface of the rotating shaft 21 and the inner peripheral surface
of the through hole 7a.
[0070] The body 3 has an intake port (intake flow path) 5, which
leads to the low-pressure side of the hydraulic chamber 4, formed
in one side surface thereof, and has a discharge port (discharge
flow path) 6, which leads to the high-pressure side of the
hydraulic chamber 4, formed in another side surface opposite said
side surface thereof. The intake port 5 and the discharge port 6
are provided so that their axes are positioned at the middle
between the rotating shafts 21, 24 of the pair of gears 20, 23.
[0071] The pair of side plates 30, 32 are plate-shaped members
having a substantially 8-shaped cross-section and respectively have
two through holes 31, 33 formed therein, they are disposed on both
sides of the gears 20, 23 in a state where the rotating shafts 21,
24 of the gears 20, 23 are inserted through the through holes 31,
33, and one end surfaces of the side plates 30, 32 are each in
contact with the entire end surfaces of the gears 20, 23 including
their tooth portions.
[0072] As shown in FIGS. 3 and 4, the bushes 40, 44 are metal
bearings comprising a member having a substantially 8-shaped
cross-section and respectively have two support holes 41, 45, and
they are respectively disposed outside the pair of side plates 30,
32 with the rotating shafts 21, 24 of the gears 20, 23 inserted
through the support holes 41, 45 and support the rotating shafts
21, 24 so that they are rotatable.
[0073] Further, dividing seals 43, 47 with elasticity, which have a
substantially FIG. 3 shape in side view, are provided on end
surfaces facing the side plates 30, 32 of the bushes 40, 44,
respectively. The dividing seals 43, 47 respectively divide gaps
50, 51 between the bushes 40, 44 and the side plates 30, 32 into a
high-pressure side and a low-pressure side, and a hydraulic oil in
the high-pressure side of the hydraulic chamber 4 is introduced
into the high-pressure sides of the gaps 50, 51 through an
appropriate flow path and the one end surfaces of the side plates
30, 32 are pressed onto the end surfaces of the gears 20, 23 by the
high-pressure hydraulic oil introduced into the gaps 50, 51,
thereby preventing leakage of the hydraulic oil from the
high-pressure side to the low-pressure side. It is noted that,
although the high-pressure hydraulic oil in the hydraulic chamber 4
acts also on end surfaces facing the gears 20, 23 of the side
plates 30, 32, the side plates 30, 32 respectively have a larger
pressure receiving area in the gaps 50, 51 than on their respective
gears 20, 23 sides, and, as a result thereof, the side plates 30,
32 are pressed onto the end surfaces of the gears 20, 23 by the
difference between the acting forces applied thereto.
[0074] Further, the other end surfaces of the bushes 40, 44 are in
contact with end surfaces of the front cover 7 and the end cover
11, respectively, thereby creating a state where the end surfaces
of the gears 20, 23 and the one end surfaces of the side plates 30,
32 are in contact with each other and the other end surfaces of the
side plates 30, 32 and the dividing seals 43, 47 provided on the
bushes 40, 44 are in contact with each other and a state where the
gears 20, 23, the side plates 30, 32 and the bushes 40, 44 are
pressurized.
[0075] Further, the intermediate plate 8 has a cylinder hole 8a
formed at a portion facing an end surface of the rotating shaft 21
on the rear side of the gear 20 thereof, and a piston 9 is inserted
through the cylinder hole 8a. The end cover 11 has a recess portion
11a formed at a portion corresponding to the cylinder hole 8a
thereof, and the hydraulic oil in the high-pressure side of the
hydraulic chamber 4 is supplied into the recess portion 11a through
a not-shown flow path, so that the hydraulic oil in the
high-pressure side acts on the back surface (rear end surface) of
the piston 9.
[0076] As described above, in this embodiment, the gear 20 has a
right-handed helical tooth portion and the gear 23 has a
left-handed helical tooth portion. Therefore, when the gear 20 is
rotated in the direction indicated by the arrow (clockwise
rotation), a backward pressure receiving thrust force F.sub.pa
generated by the high-pressure hydraulic oil acting on the tooth
portion of the gear 20 and a similarly backward meshing thrust
force F.sub.ma generated by meshing of the gears 20, 23 act on the
gear 20, and therefore a combined thrust force F.sub.x which is a
resultant force of the pressure receiving thrust force F.sub.pa and
the meshing thrust force F.sub.ma acts thereon.
[0077] The size of the cross-sectional area (pressure receiving
area) of the piston 9 of this embodiment is set so that a thrust
which almost balances the combined thrust force F.sub.x acting on
the gear 20 and eliminates the combined thrust force F.sub.x is
generated by the high-pressure hydraulic oil acting on the back
surface of the piston 9.
[0078] The pressure receiving thrust force F.sub.pa, the meshing
thrust force F.sub.ma and the combined thrust force F.sub.x can be
calculated theoretically. Hereinafter, the theoretical calculation
will be explained. It is noted that the meanings of the references
used in the explanation given below are as follows:
V.sub.th: theoretical discharge amount per revolution of pump
(gear) (m.sup.3/rev) r.sub.w: radius of working pitch circle of
gear (m) b: face width of gear (m) h: tooth depth of gear (m) Q:
discharge flow rate of pump (m.sup.3/sec) P.sub.th: hydraulic
pressure of pump not taking into account losses (Pa) P: hydraulic
pressure of pump taking into account losses (Pa) .eta..sub.m:
mechanical efficiency of pump .beta..sub.w: working helix angle of
gear (rad) .beta..sub.b: base cylinder helix angle of gear (rad)
T.sub.d: input shaft torque applied to rotating shaft of driving
gear (Nm) n: number of revolution of rotating shaft of gear
(rev/sec) .omega.: angular velocity applied to rotating shaft of
driving gear (rad/sec)=2.times..pi..times.n T.sub.m: transmitted
torque from driving gear to driven gear (Nm) W.sub.p: workload
applied to liquid by driving of pump (J=Nm) F.sub.wt: nominal
working tangential force (N) F.sub.n: tooth surface normal force
(N) F.sub.pt: transverse tooth surface normal force (N)
.alpha..sub.wt: working transverse pressure angle (rad) F.sub.ma:
meshing thrust force (N) F.sub.pa: pressure receiving thrust force
(N) F.sub.x: combined thrust force (N) .epsilon..sub..alpha.:
transverse contact ratio .epsilon..sub..beta.: overlap ratio
.epsilon..sub.r: ratio of contact ratios
(.epsilon..sub..beta./.epsilon..sub..alpha.)
[0079] [Meshing Thrust Force]
[0080] Hereinafter, calculation of the meshing thrust force
F.sub.ma will be explained.
[0081] First, in a case where mechanical efficiency .eta..sub.m is
not taken into account, the following equation holds because an
input energy (T.sub.d.times..omega.) and an output energy
(P.sub.th.times.Q) are equal to each other.
T.sub.d.times..omega.=P.sub.th.times.Q=P.sub.th.times.V.sub.th.times.n
(Equation 1)
[0082] It is noted that, in a case where the mechanical efficiency
.eta..sub.m is taken into account, the following equation
holds:
T.sub.d.times..omega.=P.sub.th.times.V.sub.th.times.n/.eta..sub.m,
and (Equation 2)
[0083] the hydraulic pressure of pump (pressure of hydraulic oil) P
taking into account the mechanical efficiency .eta..sub.m is
represented by the following equation.
P=P.sub.th.times..eta..sub.m (Equation 3)
[0084] Further, because the theoretical discharge amount of pump
V.sub.th is approximated by the theoretical discharge amount of two
gears, it can be represented by the following equation.
V.sub.th.apprxeq.2.pi..times.r.sub.w.times.h.times.b (Equation
4)
[0085] Further, on the basis of the Equation 1, the Equation 4 and
the relationship of .omega.=2.pi..times.n, the relationship between
driving torque and hydraulic pressure of the pump can be
represented by the following equation.
Td.apprxeq.2.pi..times.r.sub.w.times.h.times.b.times.P.sub.th.times.n/.o-
mega.=r.sub.w.times.h.times.b.times.P.sub.th (Equation 5)
[0086] Furthermore, because the gears of the pump have the same
geometric shape and their workloads are equal to each other, the
transmitted torque T.sub.m transmitted from the driving gear to the
driven gear can be represented by the following equation.
T.sub.m.apprxeq.0.5T.sub.d=0.5r.sub.w.times.h.times.b.times.P.sub.th
(Equation 6)
[0087] The transmitted torque T.sub.m and the nominal tangential
force generated on the working pitch circle (nominal working
tangential force) F.sub.wt have the relationship represented by the
following equation.
F.sub.wt=T.sub.m/r.sub.w (Equation 7)
[0088] Further, as shown in FIG. 5, because the nominal working
tangential force F.sub.wt is a working-pitch-circle circumferential
component of the transverse tooth surface normal force F.sub.nt
which is obtained by projecting the tooth surface normal force
F.sub.n on the transverse cross-section of the gear, the
relationship between them can be represented by the following
equations.
F.sub.wt=F.sub.nt.times.cos .alpha..sub..omega.t (Equation 8)
F.sub.nt=F.sub.n.times.cos .beta..sub.b (Equation 9)
F.sub.n=F.sub.wt/(cos .alpha..sub..omega.t.times.cos .beta..sub.b)
(Equation 10)
F.sub.ma=F.sub.n.times.sin .beta..sub.b (Equation 11)
[0089] On the basis of the Equations 8 to 11, the meshing thrust
force F can be represented by the following equation.
F.sub.ma=F.sub.wt.times.tan .beta..sub.b/cos .alpha..sub.wt
(Equation 12)
[0090] Further, on the basis of the basic theory of helical gear,
there is the relationship of
tan .beta..sub.b=tan .beta..sub.w.times.cos
.alpha..sub..omega.t,
[0091] and therefore, on the basis of this relationship and the
Equations 6, 7 and 12, eventually the meshing thrust force F.sub.ma
can be represented by the following equation.
F.sub.ma.apprxeq.0.5h.times.b.times.P.sub.th.times.tan
.beta..sub.w
[0092] The meshing thrust force F.sub.ma calculated by the Equation
13 acts on the gears 20, 23.
[0093] [Pressure Receiving Thrust Force]
[0094] In a helical gear (continuous-line-of-contact meshing gear)
which has a tooth profile, as shown in FIG. 6, including an arc
portion in a tooth tip and a tooth root and forming a continuous
line of contact (line of meshing contact) from one end to the other
end in a face width direction at a meshing portion, the line of
meshing contact separates a discharge side and an intake side, and
therefore an acting force generated by the pressure difference
between both sides of the line of contact acts on a tooth on which
the line of contact is formed, and the pressure receiving thrust
force F.sub.pa, which is a thrust-directional component along the
gear shaft of the acting force, can be evaluated by multiplying an
area obtained by projecting a tooth surface on which a hydraulic
pressure acts on a plane perpendicular to the gear shaft (rotating
shaft) (see FIG. 7) and the hydraulic pressure force.
[0095] Further, because the pressure receiving thrust force
F.sub.pa varies depending on the meshing manner of the pair of
gears, this has to be calculated in accordance with the meshing
manner. In the field of gear, as indices of the meshing manner, an
index called the transverse contact ratio .epsilon..sub..alpha. and
an index called the overlap ratio .epsilon..sub..beta. are known.
Generally the distance between teeth measured in the normal
direction of the tooth is called the normal pitch and the length of
actual meshing on the line of action is called the length of
action, and the transverse contact ratio .epsilon..sub..alpha. is
the value obtained by dividing the length of action by the normal
pitch. Further, in a case of helical gears, because their tooth
traces are helical, the length of meshing between a pair of teeth
is longer than that in a case of spur gears, and the increment of
the contact ratio due to their helices is called the overlap ratio
.epsilon..sub..beta., and when the length of the long meshing due
to their helices is evaluated on the plane of action, it is
b.times.tan .beta..sub.b, and therefore the overlap ratio
.epsilon..sub..beta. can be represented by the following
equation.
.epsilon..sub..beta.=b.times.tan .beta..sub.b/p.sub.b=b.times.tan
.beta..sub.w/p.sub.w, (Equation 14)
[0096] where p.sub.b is the normal pitch and p.sub.w is the pitch
on the pitch circle.
[0097] Further, in the present invention, the ratio of contact
ratios .epsilon..sub.r (=.epsilon..sub..beta./.epsilon..sub..beta.)
which is the ratio of the transverse contact ratio
.epsilon..sub..alpha. to the overlap ratio .epsilon..sub..beta. is
used as an index of the meshing manner of the helical gears. The
reason therefor is that, because, in a case of a
"continuous-line-of-contact meshing gear", the state of a line of
contact at a meshing portion varies depending on the value of the
ratio of contact ratios .epsilon..sub.r and therefore an area where
a hydraulic pressure acts on the tooth surface varies, it is
necessary to perform case classification based on the value of the
ratio of contact ratios .epsilon..sub.r and evaluate the area where
a hydraulic pressure acts on the tooth surface to calculate the
pressure receiving thrust force F.sub.pa which is generated by the
hydraulic pressure.
[0098] It is noted that, as for what kind of line of contact is
formed in accordance with the value of the ratio of contact ratios
.epsilon..sub.r, specific modes are shown in FIGS. 8 to 11. The
example shown in FIG. 8 is a case of 1<.epsilon..sub.r<2, the
example shown in FIG. 9 is a case of .epsilon..sub.r=2, the example
shown in FIG. 10 is a case of 2<.epsilon..sub.r<3, and the
example shown in FIG. 11 is case of .epsilon..sub.r=3. In the
examples shown in FIGS. 8 and 9, a line of contact is formed on one
tooth when one end of the line of contact is located at a tooth
root, and, in the examples shown in FIGS. 10 and 11, a line of
contact is formed across two teeth when one end of the line of
contact is similarly located at a tooth root.
[0099] Next, a method of calculating the area where a hydraulic
pressure acts on a tooth surface of a gear is explained.
[0100] FIGS. 12 and 13 show plan views showing a meshing portion of
gears, and FIG. 12 shows gears having a tooth profile which
provides a ratio of contact ratios .epsilon..sub.r in the range of
1<=.epsilon..sub.r<=2, and FIG. 13 shows gears having a tooth
profile which provides a ratio of contact ratios .epsilon..sub.r in
the range of 2<=.epsilon..sub.r<=3. In each figure, the
oblique solid lines indicate ridge lines of tooth tips and the
oblique broken lines indicate lines of tooth roots.
[0101] First, in a case of gears having a tooth profile which
provides a ratio of contact ratios .epsilon..sub.r in the range of
1<=.epsilon..sub.r<=2, a hydraulic pressure acts on regions
a.sub.1, a.sub.2 and a.sub.3 with a line of meshing contact L as a
border. The hydraulic pressure acts on the regions a.sub.1 and
a.sub.3 in the same thrust direction, and the hydraulic pressure
acts on the region a.sub.2 in the opposite thrust direction.
Therefore, an effective pressure receiving area Ap.sub.1 taking
into account a cancellation by the difference of direction can be
represented by the following equation, wherein the area from tooth
root to tooth tip of one tooth surface is A.
Ap.sub.1=A((.epsilon..sub.r-1).sup.2+1)/2.epsilon..sub.r (Equation
15)
[0102] Similarly, in a case of gears having a tooth profile which
provides a ratio of contact ratios .epsilon..sub.r in the range of
2<=.epsilon..sub.r<=3, because a hydraulic pressure acts on
regions a.sub.4 and a.sub.6 in the same thrust direction and acts
on a region a.sub.5 in the opposite thrust direction with a line of
meshing contact L as a border, an effective pressure receiving area
Ap.sub.2 taking into account a cancellation by the difference of
direction can be represented by the following equation.
Ap.sub.2=A-A((.epsilon..sub.r-2).sup.2+2)/2.epsilon..sub.r
(Equation 16)
[0103] As described above, the effective pressure receiving area
which causes a thrust force due to a hydraulic pressure varies
depending on the value of the ratio of contact ratios
.epsilon..sub.r.
[0104] Next, the pressure receiving thrust force F.sub.pa is
calculated on the basis of the pressure receiving area Ap.sub.1,
Ap.sub.2 obtained in the way as described above. It is noted that
an area A.sub.x obtained by projecting the area A on a plane
perpendicular to the gear shaft can be evaluated by the following
equation on the basis of an angle of rotation .theta. of a tooth
seen from the plane perpendicular to the gear shaft, the radius of
working pitch circle r.sub.w and the tooth depth h.
A.sub.x=h.times.r.sub.w.times..theta.=h.times.b.times.tan
.beta..sub.w (Equation 17)
[0105] [Pressure Receiving Thrust Force not Taking into Account
Mechanical Efficiency]
[0106] As described above, the pressure receiving thrust force
F.sub.pa can be evaluated by multiplying an area obtained by
projecting a tooth surface on which a hydraulic pressure acts on a
plane perpendicular to the gear shaft (rotating shaft), that is,
the area A.sub.x and the hydraulic pressure force.
[0107] Therefore, in the case of 1<=.epsilon..sub.r<=2, the
pressure receiving thrust force F.sub.pa1 which is generated by the
hydraulic pressure P.sub.th not taking into account the mechanical
efficiency .eta..sub.m can be represented by the following equation
on the basis of the Equations 15 and 17.
F pa 1 = P th .times. Ap 1 = P th .times. h .times. b .times. tan
.beta. w .times. ( ( r - 1 ) 2 + 1 ) / 2 r ( Equation 18 )
##EQU00001##
[0108] Further, in the case of 2<=.epsilon..sub.r<=3, the
pressure receiving thrust force F.sub.pa2 which is generated by a
hydraulic pressure P.sub.th not taking into account the mechanical
efficiency .eta..sub.m can be represented by the following equation
on the basis of the Equations 16 and 17.
F pa 2 = P th .times. Ap 2 = P th .times. h .times. b .times. tan
.beta. w .times. ( 2 r - ( ( r - 2 ) 2 + 2 ) ) / 2 r ( Equation 19
) ##EQU00002##
[0109] [Combined Thrust Force not Taking into Account Mechanical
Efficiency]
[0110] On the basis of the above-described Equations 13, 18 and 19,
in a case of the oil hydraulic pump 1 as shown in FIG. 1, the
combined thrust force F.sub.xp acting on the driving gear 20 and
the rotating shaft 21 can be represented by the following
equation.
in the case of 1 <= r <= 2 F xp 1 = F ma + F pa 1 .apprxeq.
0.5 h .times. b .times. P th .times. tan .beta. w + P th .times. h
.times. b .times. tan .beta. w .times. ( ( r - 1 ) 2 + 1 ) / 2 r (
Equation 20 ) in the case of 2 <= r <= 3 F xp 2 = F ma + F pa
2 .apprxeq. 0.5 h .times. b .times. P th .times. tan .beta. w + P
th .times. h .times. b .times. tan .beta. w .times. ( 2 r - ( ( r -
2 ) 2 + 2 ) ) / 2 r ( Equation 21 ) ##EQU00003##
[0111] On the other hand, the combined thrust force F.sub.xg acting
on the driven gear 23 and the rotating shaft 24 can be represented
by the following equation.
in the case of 1 <= r <= 2 F xg 1 = - F ma + F pa 1 .apprxeq.
- 0.5 h .times. b .times. P th .times. tan .beta. w + P th .times.
h .times. b .times. tan .beta. w .times. ( ( r - 1 ) 2 + 1 ) / 2 r
( Equation 22 ) in the case of 2 <= r <= 3 F xg 2 = - F ma +
F pa 2 .apprxeq. - 0.5 h .times. b .times. P th .times. tan .beta.
w + P th .times. h .times. b .times. tan .beta. w .times. ( 2 r - (
( r - 2 ) 2 + 2 ) / 2 r ) ( Equation 23 ) ##EQU00004##
[0112] Further, on the basis of the Equations 20 to 23, when the
ratio of contact ratios .epsilon..sub.r is set to 1, 2 or 3, the
combined thrust forces F.sub.xp and F.sub.xg are as follows. It is
noted that the combined thrust forces when .epsilon..sub.r=1 are
F.sub.xp1' and F.sub.xg1', the combined thrust forces when
.epsilon..sub.r=2 are F.sub.xp2' and F.sub.xg2', and the combined
thrust forces when .epsilon..sub.r=3 are F.sub.xp3' and
F.sub.xg3'.
F.sub.xp1'.apprxeq.h.times.b.times.P.sub.th.times.tan .beta..sub.w
(Equation 24)
F.sub.xg1'.apprxeq.-0.5h.times.b.times.P.sub.th.times.tan
.beta..sub.w+(P.sub.th.times.h.times.b.times.tan .beta..sub.w)/2=0
(Equation 25)
F.sub.xp2'.apprxeq.h.times.b.times.P.sub.th.times.tan .beta..sub.w
(Equation 26)
F.sub.xg2'.apprxeq.-0.5h.times.b.times.P.sub.th.times.tan
.beta..sub.w+(P.sub.th.times.h.times.b.times.tan .beta..sub.w)/2=0
(Equation 27)
F.sub.xp3'.apprxeq.h.times.b.times.P.sub.th.times.tan .beta..sub.w
(Equation 28)
F.sub.xg3'.apprxeq.-0.5h.times.b.times.P.sub.th.times.tan
.beta..sub.w+(P.sub.th.times.h.times.b.times.tan .beta..sub.w)/2=0
(Equation 29)
[0113] Thus, in a case where it is assumed that mechanical losses
are not taken into account, that is, the mechanical efficiency
.eta..sub.m is 100%, when the ratio of contact ratios
.epsilon..sub.r is set to 1, 2 or 3, the combined thrust force
F.sub.xg1', F.sub.xg2', F.sub.xg3' acting on the driven gear 23 and
the rotating shaft 24 is 0 in each case, and it is seen that the
driven gear 23 and the rotating shaft 24 are in a state where no
thrust force acts thereon. On the other hand, the combined thrust
force F.sub.xp1', F.sub.xp2', F.sub.xp3' acting on the driving gear
20 and the rotating shaft 21 is h.times.b.times.P.sub.th.times.tan
.beta..sub.w in each case.
[0114] In view of the foregoing, in the case where mechanical
losses are not taken into account, setting the ratio of contact
ratios .epsilon..sub.r to 1, 2 or 3 makes it possible to create a
state where no thrust force acts on the driven gear 23 and the
rotating shaft 24, and applying a force equal to
h.times.b.times.P.sub.th.times.tan .beta..sub.w to the rotating
shaft 21 of the driving gear 20 as a drag makes it possible to
create a state where no thrust force acts on the driving gear 20,
the rotating shaft 21, the driven gear 23 and the rotating shaft
24. It is noted that, in a case of .epsilon..sub.r<=1, it is not
possible to obtain practical gears 20, 23.
[0115] Thus, in an oil hydraulic pump (hydraulic device) using
"continuous-line-of-contact meshing gears", in a case where
mechanical losses are not taken into account, setting the tooth
profiles of the driving gear 20 and the driven gear 23 to such a
tooth profile that the ratio of contact ratios .epsilon..sub.r is 2
or 3 makes it possible to create a state where no thrust force acts
on the driven gear 23 and the rotating shaft 24. However, because a
hydraulic device always involves mechanical losses, in the strict
sense, it is required that no thrust force act on the driven gear
23 and the rotating shaft 24 in a state where the mechanical
efficiency .eta..sub.m is taken into account. Therefore,
hereinafter, the combined thrust forces F.sub.xp, F.sub.xg taking
into account the mechanical efficiency n, are considered.
[0116] [Pressure Receiving Thrust Force Taking into Account
Mechanical Efficiency]
[0117] The pressure receiving thrust force F.sub.pa1 generated by
the hydraulic pressure P taking into account the mechanical
efficiency .eta..sub.m can be represented by the following equation
which is made by replacing P.sub.th in the Equations 18 and 19 with
P.
in the case of 1<=.epsilon..sub.r<=2
F.sub.pa1=P.times.h.times.b.times.tan
.beta..sub.w.times.((.epsilon..sub.r-1).sup.2+1)/2.epsilon..sub.r
(Equation 30)
in the case of 2<=.epsilon..sub.r<=3
F.sub.pa2=P.times.h.times.b.times.tan
.beta..sub.w.times.(2.epsilon..sub.r-((.epsilon..sub.r-2).sup.2+2))/2.eps-
ilon..sub.r (Equation 31)
[0118] [Combined Thrust Force Taking into Account Mechanical
Efficiency]
[0119] Further, the combined thrust force F.sub.xp acting on the
driving gear 20 and the rotating shaft 21 and the combined thrust
force F.sub.xg acting on the driven gear 23 and the rotating shaft
24, which are combined thrust forces taking into account the
mechanical efficiency .eta..sub.m, are represented by the following
equations.
in the case of 1<=.epsilon..sub.r<=2
F.sub.xp1.apprxeq.0.5h.times.b.times.P.sub.th.times.tan
.beta..sub.w+P.times.h.times.b.times.tan
.beta..sub.w.times.((.epsilon..sub.r-1).sup.2+1)/2.epsilon..sub.r
(Equation 32)
in the case of 2<=.epsilon..sub.r<=3
F.sub.xp2.apprxeq.0.5h.times.b.times.P.sub.th.times.tan
.beta..sub.w+P.times.h.times.b.times.tan
.beta..sub.w.times.(2.epsilon..sub.r-((.epsilon..sub.r-2).sup.2+2))/2.eps-
ilon..sub.r (Equation 33)
in the case of 1<=.epsilon..sub.r<=2
F.sub.xg1.epsilon.-0.5h.times.b.times.P.sub.th.times.tan
.beta..sub.w+P.times.h.times.b.times.tan
.beta..sub.w.times.((.epsilon..sub.r-1).sup.2+1)/2.epsilon..sub.r
(Equation 34)
in the case of 2<=.epsilon..sub.r<=3
F.sub.xg2.apprxeq.-0.5h.times.b.times.P.sub.th.times.tan
.beta..sub.w+P.times.h.times.b.times.tan
.beta..sub.w.times.(2.epsilon..sub.r-((.epsilon..sub.r-2).sup.2+2)/2.epsi-
lon..sub.r) (Equation 35)
[0120] In view of the foregoing, although the inventors considered,
using the Equations 34 and 35, a case where the combined thrust
force F.sub.xg2 acting on the driven gear 23 and the rotating shaft
24 would be 0, a practical solution could not be obtained in the
case of 1<=.epsilon..sub.r<=2. On the other hand, they found
out that a practical solution could be obtained in the case of
2<=.epsilon..sub.r<=3.
[0121] Although it is said that a practical range of the mechanical
efficiency n, is generally 0.91<=.eta..sub.m<=0.99, if
.eta..sub.m=0.95, .epsilon..sub.r which makes F.sub.xg2 0 in the
Equation 35 is calculated by the following equation. It is noted
that P=P.sub.th.times..eta..sub.m holds on the basis of the
Equation 3.
0.5P.sub.th.times.h.times.b.times.tan
.beta..sub.w=0.95P.sub.th.times.h.times.b.times.tan
.beta..sub.w.times.(2.epsilon..sub.r-((.epsilon..sub.r-2).sup.2.+-.2)/2.e-
psilon..sub.r)0.5/0.95=(2.epsilon..sub.r-((.epsilon..sub.r-2).sup.2+2)/2.e-
psilon..sub.r) (Equation 36)
[0122] When solving the quadratic equation of the Equation 36, two
solutions, .epsilon..sub.r=2.13, 2.82, are obtained. Therefore, in
a case where it is assumed that the mechanical efficiency
.eta..sub.m=0.95, the combined thrust force F.sub.xg2 acting on the
driven gear 23 and the rotating shaft 24 can be made 0 by making
the gears to have such a tooth profile that the ratio of contact
ratios .epsilon..sub.r is 2.13 or 2.82.
[0123] Taking into consideration the foregoing, when evaluating the
relationship between .epsilon..sub.r and .eta..sub.m which makes
F.sub.xg2 0 in the Equation 35, the following equation holds.
0.5 P th .times. h .times. b .times. tan .beta. w = .eta. m .times.
P th .times. h .times. b .times. tan .beta. w .times. ( 2 r - ( ( r
- 2 ) 2 + 2 ) / 2 r ) .eta. m = 2 r / ( 2 .times. ( 2 r - ( ( r - 2
) 2 + 2 ) ) ) = r / ( 6 r - r 2 - 6 ) ( Equation 37 )
##EQU00005##
[0124] Thus, by calculating, using the Equation 37, a ratio of
contact ratios .epsilon..sub.r which meets the Equation 37 in
accordance with a mechanical efficiency .eta..sub.m which is
assumed to be preferable for practical use and making the gears 20,
23 to have a tooth profile corresponding to the calculated ratio of
contact ratios .epsilon..sub.r, the combined thrust force F.sub.xg2
acting on the driven gear 23 and the rotating shaft 24 can be made
0.
[0125] As described above, by making the gears 20, 23 to have such
a tooth profile that the ratio of contact ratios .epsilon..sub.r
meets 2<=.epsilon..sub.r<=3, the combined thrust force
F.sub.xg acting on the driven gear 23 and the rotating shaft 24 can
be made 0 within an appropriate mechanical efficiency .eta..sub.m.
That is, it is possible to create a state where no thrust force
acts on the driven gear 23 and the rotating shaft 24. Further, in
this embodiment, the gears 20, 23 have such a tooth profile.
[0126] On the other hand, in a case where the gears 20, 23 are made
to have such a tooth profile that the ratio of contact ratios
.epsilon..sub.r meets 2<=.epsilon..sub.r<=3, the combined
thrust force F.sub.xp (=F.sub.xp2) calculated by the Equation 33
acts on the driving gear 20 and the rotating shaft 21. Therefore,
when a thrust of the piston 9 pressing the rotating shaft 21 is
equal to the combined thrust force F.sub.xp calculated by the
Equation 33, they are balanced and a state where no thrust force
acts on the rotating shaft 21 can be created. Further, for causing
the piston 9 to generate such a thrust, the cross-sectional area S
(mm.sup.2) of the piston 9 can be calculated by the following
equation, where the pressure of the hydraulic oil in the high
pressure side is P (the pressure of the hydraulic oil taking into
account the mechanical efficiency).
S.times.P=F.sub.xp(=F.sub.xp2)
S.times.P=0.5h.times.b.times.P.times.tan
.beta..sub.w/.eta..sub.m+P.times.h.times.b.times.tan
.beta..sub.w.times.(2.epsilon..sub.r-((.epsilon..sub.r-2).sup.2+2))/2.eps-
ilon..sub.r
S=0.5h.times.b.times.tan
.beta..sub.w/.eta..sub.m+h.times.b.times.tan
.beta..sub.w.times.(2.epsilon..sub.r-((.epsilon..sub.r-2).sup.2+2))/2.eps-
ilon..sub.r (Equation 38)
[0127] It is noted that, because the oil hydraulic pump 1 involves
various variable elements such as variation in machining and
assembling and variation related to the modulus of elasticity of an
elastic seal for enabling the rotating shafts to move in their
axial directions and the combined thrust force F.sub.xp also varies
in accordance with the variable elements, taking this into
consideration, it is preferred that the cross-sectional area S is
set to meet the following equation.
0.9(F.sub.xp/P)<=S<=1.1(F.sub.xp/P) (Equation 39)
[0128] According to the oil hydraulic device 1 having the
above-described configuration, appropriate piping which is
connected to an appropriate tank for storing a hydraulic oil
therein is connected to the intake port 5 of the housing 2 and
appropriate piping which is connected to an appropriate oil
hydraulic equipment is connected to the discharge port 6, and
further an appropriate drive motor is connected to the screw
portion 22 of the rotating shaft 21 of the driving gear 20. Then,
the drive motor is driven to rotate the driving gear 20.
[0129] Thereby, the driven gear 23 meshing with the driving gear 20
rotates, a hydraulic oil in a space between the inner peripheral
surface 3a of the hydraulic chamber 4 and the tooth portions of the
gears 20, 23 is transferred to the discharge port 6 side by the
rotations of the gears 20, 23, and thereby the discharge port 6
side becomes a high-pressure side and the intake port 5 side
becomes a low-pressure side with the meshing portion of the pair of
gears 20, 23 as a border.
[0130] Further, when the intake port 5 side is brought into a
negative pressure by the transfer of the hydraulic oil to the
discharge port 6 side, the hydraulic oil in the tank is inhaled
into the low-pressure side of the hydraulic chamber 4 through the
piping and the intake port 5, and, similarly, the hydraulic oil in
the space between the inner peripheral surface of the hydraulic
chamber 4 and the tooth portions of the gears 20, 23 is transferred
to the discharge port 6 side by the rotations of the gears 20, 23
and is pressurized to a high pressure and transmitted to the oil
hydraulic equipment through the discharge port 6 and the
piping.
[0131] Further, the high-pressure hydraulic oil is lead into the
gaps 50, 51 between the bushes 40, 44 and the side plates 30, 32
through the flow path and the side plates 30, 32 are pressed onto
the end surfaces of the gears 20, 23 by the function of the
hydraulic oil, thereby preventing leakage of the hydraulic oil from
the high-pressure side to the low-pressure side.
[0132] By the way, as described above, in the oil hydraulic pump 1
using the helical gears 20, 23 of this embodiment, although the
combined thrust force F.sub.x, which is a resultant force of the
pressure receiving thrust force F.sub.pa and the meshing thrust
force F.sub.ma, acts on the gear 20, since a force which almost
balances and resists the combined thrust force F.sub.x is caused to
act on the rear end surface of the rotating shaft 21 of the gear 20
by the piston 9, a state where no thrust force acts on the gear 20
is achieved.
[0133] On the other hand, since the pressure receiving thrust force
F.sub.pa and the meshing thrust force F.sub.ma act on the gear 23
in the opposite directions, they are cancelled, and, particularly,
using "continuous-line-of-contact meshing gears" as the helical
gears 20, 23 like this embodiment and making the gears to have such
a tooth profile that the ratio of contact ratios .epsilon..sub.r
meets 2<=.epsilon..sub.r<=3 makes it possible to create a
state where no thrust force acts on the gear 23.
[0134] Thus, in the oil hydraulic pump 1 of this embodiment, a
state where both of the pair of gears 20, 23 do not receive a
thrust-directional force can be achieved, and therefore the
above-described conventional problem that seizure or damage due to
a thrust force occurs on the side plates 30, 32 which are in
sliding contact with both end surfaces of the pair of gears 20, 23
is not caused.
[0135] Further, since the hydraulic oil in the high-pressure side
is caused to act on the back surfaces of the side plates 30, 32 and
thereby the side plates 30, 32 are brought into tight contact with
both end surfaces of the gears 20, 23, and the side plates 30, 32
are supported by bringing the dividing seals 43, 47 with elasticity
into tight contact with the back surfaces of the side plates 30,
32, even if periodic variation occurs on the pressure receiving
thrust force F.sub.pa or the meshing thrust force F.sub.ma or
sudden vibration occurs on the oil hydraulic pump 1, such variation
and sudden vibration are absorbed by movement of the gears 20, 23
and the side plates 30, 32 in the axial directions of the rotating
shafts 21, 24 by elastic deformation of the dividing seals 43, 47,
thereby suppressing the occurrence of noise caused by such
variation and vibration.
[0136] Further, in the oil hydraulic pump 1 of this embodiment,
since providing the piston 9 for causing a reaction force to act on
only the rotating shaft 21 of the gear 20 achieves the state where
no thrust force acts on both of the pair of gears 20, 23, it is
possible to solve the above-described conventional problem while
reduing costs for manufacturing the oil hydraulic pump 1.
[0137] Thus, although one embodiment of the present invention has
been described, a specific mode in which the present invention can
be realized is not limited thereto.
[0138] For example, although the above-described embodiment has the
configuration in which the side plates 30, 32 are provided between
the gears 20, 23 and the bushes 40, 44 to be in contact with the
gears 20, 23 and the spaces between the bushes 40, 44 and the side
plates 30, 32 are divided by the dividing seals 43, 47, the present
invention includes also modes in which the side plates 30, 32 and
the dividing seals 43, 47 as described above are not provided.
[0139] Further, in a mode in which the side plates 30, 32 are not
provided, as shown in FIGS. 14 and 15, there may be an oil
hydraulic pump 1' having a configuration in which bushes 40', 44'
are disposed to be in contact with the end surfaces of the gears
20, 23, a diving seal 43' with elasticity is interposed between the
bush 40' and the front cover 7 and a diving seal 47' with
elasticity is interposed between the bush 44' and the intermediate
cover 8, and a high oil pressure is supplied into a space 50'
between the bush 40' and the front cover 7 and a space 51' between
the bush 44' and the intermediate cover 8.
[0140] Also in this configuration, the bushes 40', 44' are pressed
onto the end surfaces of the gears 20, 23, thereby preventing
leakage of the hydraulic oil through the end surfaces of the gears
20, 23. Further, the movability of the gears 20, 23 and the bushes
40', 44' in the axial directions of the rotating shafts 21, 24 is
secured by elastic deformation of the dividing seals 43', 47', and
even if periodic variation occurs on the pressure receiving thrust
force F.sub.pa or the meshing thrust force F.sub.ma or sudden
vibration occurs on the oil hydraulic pump 1', these are absorbed
by the movement of the gears 20, 23 and the bushes 40', 44' in the
axial directions, thereby suppressing the occurrence of noise
caused by the variation and the vibration.
[0141] It is noted that, in FIG. 14, the same components as those
of the oil hydraulic pump 1 shown in FIGS. 1 to 4 are indicated by
the same references.
[0142] Further, although, in the oil hydraulic pump 1 of the
above-described embodiment, a right-handed helical gear is used as
the driving gear 20 and a left-handed helical gear is used as the
driven gear 23, there may be an oil hydraulic pump 1'' using a
left-handed helical gear as a driving gear 20'' and a right-handed
helical gear as a driven gear 23'', as shown in FIG. 16. In this
case, the driving gear 20'' is rotated in the direction indicated
by the arrow in FIG. 16.
[0143] Also in the oil hydraulic pump 1'' having this
configuration, a state where both of the gears 20'', 23'' do not
receive a thrust-directional force can be achieved and the
conventional problem that seizure or damage due to a thrust force
occurs on the side plates 30, 32 which are in sliding contact with
the gears 20'', 23 is not caused.
[0144] It is noted that, also in FIG. 16, the same components as
those of the oil hydraulic pump 1 shown in FIGS. 1 to 4 are
indicated by the same references.
[0145] Further, although in the foregoing, the embodiment in which
the hydraulic device of the present invention is embodied as an oil
hydraulic pump is shown as an example, the hydraulic device of the
present invention is not limited thereto and may be embodied as an
oil hydraulic motor, for example. Further, the working liquid is
not limited to a hydraulic oil and coolant may be used as the
working liquid, for example. In this case, the hydraulic device of
the present invention is embodied as a coolant pump.
[0146] Further, although not particularly mentioned in the
foregoing, a configuration is possible in which a key groove is
formed in the tapered portion of the rotating shaft 21 and a key is
inserted into the key groove and an appropriate rotary body may be
coupled to the tapered portion of the rotating shaft 21 by the key
groove and the key.
[0147] Further, although, in the above embodiment, the intake port
5 and the discharge port 6 are formed as through holes on the body
3, they may be anything as long as they lead to the hydraulic
chamber 4, and therefore, the intake port 5 and the discharge port
6 may be formed on the body and the front cover 7 and/or the end
cover 11 to form flow paths (an intake flow path and a discharge
flow path) one ends of which lead to the hydraulic chamber 4 though
an opening formed in the body 3 and the other ends of which lead to
the outside through an opening formed in the front cover 7 and/or
the end cover 11.
[0148] Furthermore the "continuous-line-of-contact meshing gear"
includes an involute gear, a sine-curve gear, a segmental gear, a
parabola gear, etc.
REFERENCE SIGNS LIST
[0149] 1 Oil hydraulic pump [0150] 2 Housing [0151] 3 Body [0152] 4
Hydraulic chamber [0153] 7 Front cover [0154] 8 Intermediate cover
[0155] 8a Cylinder hole [0156] 9 Piston [0157] 11 End cover [0158]
11a Recess portion [0159] 20 Driving gear [0160] 21 Rotating shaft
[0161] 23 Driven gear [0162] 24 Rotating shaft [0163] 30, 32 Side
plate [0164] 40, 44 Bush [0165] 43, 47 Dividing seal [0166] 50, 51
Gap
* * * * *