Note: Descriptions are shown in the official language in which they were submitted.
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Variable-volume internal gear pump
The invention relates to an internal gear pump, in
particular for use as an engine lubrication pump for
automobiles, according to the precharacterizing clause
of Claim 1.
Owing to the requirement for power losses which are as
small as possible over the total speed and power range
of the engine, engine designers are increasingly being
faced with the requirement that the oil transport of
the pump should no longer increase with the engine
speed, as in the past. The lubricating oil requirement
curve of the engine has a digressive characteristic
over the variation of the engine speed. This means
that the engine does. not have a lubricating oil
requirement which is proportional to the speed. Thus,
this is substantially smaller at high speed.
In the case of pumps which cannot be controlled, the
lubricating oil flows via a bypass valve back into the
suction side of the. pump from a certain maximum oil
pressure. This gives rise to a considerable loss of.
hydrostatic power, which is substantially avoidable in
a demand-oriented manner by an automatically controlled
pump. In fact, this loss develops exponentially with
constant specific delivery of the pump since, in
addition to the delivery, the oil pressure too
simultaneously increases over the speed. This state of
affairs has yet another aspect: when the engine is cold
and hence the lubricating oil is extremely viscous, the
oil= pressure assumes inadmissibly high values in spite
of the bypass valve, so that especially the main flow
filter of the engine is endangered. Moreover, the
ageing of the oil increases if it is subjected to
extreme stress for a long time by the narrow gaps of
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the system and with a high pressure difference at high
sheer rate. Engine lubricating pumps controllable in
their specific delivery, with more or less undesired
secondary phenomena, are already known.
A known variable-volume internal gear pump controls the
specific delivery by rotating the centre distance line
of the gear set relative to the suction and pressure
chambers in the pump housing in an oil pressure-
dependent manner. However, this has two substantial
disadvantages, namely that, in the case of a controlled
pump, unavoidable squeezing losses arise through this
so-called differential control and the pump therefore
develops considerable noises at high speed. Moreover,
the squeezing losses reduce the mechanical efficiency
of the pump in this operating range. In addition, the
squeezing losses give rise to considerable pressure
peaks between the teeth, with the result that the
components are additionally loaded and the lifetime is
reduced.
Another solution is known as a type of vane pump or
"oscillating vane pump"., in which the eccentricity of
the rotor occupied by vanes is changed relative to the
housing ring in a known manner. This has a relatively
large number of very fine components which are
fracture-sensitive and expensive to produce.
Finally, an external gear pump is also known in which
the effective tooth width of the pump is reduced with
increasing pressure by axial displacement of the two
gears relative to one another. Since these gears must
be relatively broad, the - pump housing with its
spectacle-like inner contour must also have a
corresponding length. This leads to high manufacturing
costs for machining of the housing cavern. In
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addition, external gear pumps are sensitive to
cavitation and hence noise owing to their high delivery
pulsation and owing to their radial filling on the
suction side.
It is the object of the invention to provide a pump
which is controllable in its specific delivery, in
particular an engine lubricating pump, which avoids
these disadvantages in a comprehensive manner.
This object is achieved by realizing the features of
the independent claim. Features which further develop
the invention in an alternative or advantageous manner
are described in the dependent patent claims.
The invention comprises a variable-volume internal gear
pump. The advantage of an internal gear pump over
other engine lubricating pumps controllable in their
specific delivery is in particular that firstly an
internal gear pump is superior to the external gear
pump with regard to the noise, owing to its low
instantaneous delivery pulsation over the angle of
rotation of the gears. Moreover, it can be designed
with small numbers of teeth and simultaneously an
extremely centric design. Both lead to low tooth
engagement frequency and to low hydraulic pressure
pulsations. Possible large eccentricity of the rotor
set gives rise to very large-volume delivery cells
which, at the required displacement volume, lead to
small radial dimensions of the pump. At the same time,
the exact internal machining of the pump housing is
very simple because _in principle only circular
manufacturing operations easily implementable on the
lathe are required.
The variable-volume internal gear pump according to the
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invention comprises a housing and a rotor set chamber
which is formed in the housing and which has a low
pressure chamber with an inlet opening and a high
pressure chamber with an outlet opening for a fluid.
An inner rotor held in the rotor chamber is rotatable
about an axis of rotation and can be driven by a shaft.
in the rotor set chamber, an outer rotor having an
outer rotor axis of rotation arranged eccentrically
relative to the axis of rotation is held in a rotatable
manner. The inner rotor has outer teeth and the outer
rotor has inner teeth such that the outer rotor can
rotate with the inner rotor by means of the outer-inner
teeth in a constant rotational ratio to one another
and, in the case of a rotary drive, forms delivery
cells in which the fluid is transported from the low
pressure chamber to the high pressure chamber.
According to the invention, an adjusting member which
produces an axial movement of the inner rotor is
provided. The adjusting member is guided in an axially
moveable manner in the inner teeth of the outer rotor.
The outer rotor has radial channels in the tooth spaces
between the teeth of its inner teeth in the region of
the low pressure chamber and of the high pressure
chamber. The axial position of the inner rotor
relative to the outer rotor is variable by the axial
movement of the adjusting member so that the volume of
the delivery cells can be adjusted thereby and an
internal gear pump controllable in its specific
delivery is provided.
In particular, the displacement volume and the specific
delivery of the internal gear pump are pressure
dependent, the volume of the delivery cells and hence
also the specific delivery decreasing with increasing
pressure at the outlet opening and hence with
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increasing pressure in the high pressure chamber.
The outer teeth of the inner rotor have a shape such
that axially effective springs, in particular coil
springs, can be installed between the shaft driving the
inner rotor and the tooth contour of the outer teeth
and are arranged there.
Preferably, the springs which act axially on the
adjusting member are arranged in the inner rotor
between the shaft driving the inner rotor and the tooth
contour of the outer teeth. An adjustment space which
is connected to the high pressure chamber and axially
bounded by the adjusting member is formed within the
inner teeth of the outer rotor so that a pressure of
the fluid within the adjustment space acts axially on
the adjusting member against the spring force of the
springs. The opposite arrangement of the adjustment
space, which is connected to the high pressure chamber
and hence also to the outlet opening and whose fluid
pressure acts from one side on the adjusting member,
and the spring force which presses from the other side
onto the adjusting member, the delivery cells being
present in between, have the result that, with
increasing pressure in the high pressure chamber, the
adjusting member is displaced against the spring force
of the springs and the volume of the delivery cells
decreases.
These springs are supported in the axial direction on
the shaft, for example via a pot-like intermediate
member and a securing ring. In particular, three
springs, in particular coil springs, are arranged.
uniformly distributed on the circumference on the inner
rotor.
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The commutation of the change of oil flow from the low
pressure chamber - also referred to as suction chamber
- to the high pressure chamber - also referred to as
pressure chamber - and back from the pressure chamber
to the suction chamber is effected in a gentle manner
via the outer rotor with its radial channels
communicating with the housing, so that any squeezing
of the fluid here, in particular of the oil, in the
delivery cells is substantially avoided. The
possibility of the exactly circular bore in the pump
housing for the mounting of the outer rotor and the
separation webs between the suction chamber and
pressure chamber, which are preferably to be complied
with as accurately as possible, permit this precise
commutation.
In a further development of the invention, the
adjusting member has outer teeth which fit the inner
teeth of the outer rotor with sufficient but small play
and are therefore axially displaceable therein while
providing a seal.
The geometrical shape of the inner-outer teeth, i.e. of
the outer teeth of the inner rotor and of the inner
teeth of the outer rotor which are coordinated with
them, is, for example, in the form of epicycloidal or
arc-like outer teeth on the inner rotor, which are
produced by a self-generating milling movement of the
inner teeth of the outer rotor with one tooth more.
Here, the outer teeth of the inner rotor therefore have
one tooth less than the inner teeth of the outer rotor.
This. self-generating principle, also referred to_ as
self-generating milling,-.in.which- a master -profile is.
milled in a counter-wheel, the eccentricity and the
rotational ratio being retained, is known from the
teaching on tooth systems and need not be explained in
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more detail here. The person skilled in the art is
aware that other geometrical configurations, as are
known from the prior art in the case of gear pumps, in
particular epicycloidal outer teeth, are possible. For
example, the geometrical shape of the outer-inner teeth
can be determined by epi-and hypocycloids.
The outer teeth of the inner rotor have, for example,
between 5 and 8 teeth, in particular 6 teeth.
In one embodiment of the invention, the inner rotor is
arranged in an axially displaceable and nonrotatable
manner substantially runout-free on the shaft driving
it. The arrangement in a nonrotable and axially
displaceable manner is effected, for example, by means
of a feather key.
The delivery cells are preferably closed in the axial
direction and in a position opposite to the adjusting
member by a pinion plate whose inner teeth fit the
outer teeth of the inner rotor with sufficient but
little play in such a way that the inner rotor is
axially moveable within the inner teeth of the pinion
plate. Blading corresponding to a centrifugal pump is
preferably arranged or formed on the pinion plate.
This blading corresponding to a centrifugal pump is in
particular axial blading. As a result of this blading,
the pump is so to speak pitched on the suction side so
that, with increasing speed, the liquid pressure in the
.suction chamber increases approximately with the square
of the speed. According to the invention, this pump is
therefore suitable for extremely high speeds, owing to
the avoidance of cavitation bubbles in the oil. This
too leads to small space requirement of the pump in the
engine.
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A compensating pressure region which is subjected to
high pressure, compensates the axial forces and acts as
a compensating surface can be provided on the drive
side between the housing and a drive wheel arranged
outside the housing on the shaft. Alternatively, this
compensating pressure region subjected to high pressure
and compensating the axial forces is provided on the
side opposite the drive side, between the housing or a
cover of the housing and the pot-like intermediate
member. In both cases, the compensating pressure
region, which is connected to the high pressure region
or the high pressure chamber, ensures that a hydraulic
compressive force in the compensating pressure region
counteracts the hydraulic compressive force in the
adjustment space which is likewise subjected to high
pressure.
In the drawings, the subject matter of the invention is
shown schematically, purely by way of example, with
reference to specific working examples.
Figure 1 shows an embodiment of the internal gear pump
in a longitudinal section through the middle
of the shaft and the middle of the inner
rotor of the pump at maximum specific
delivery;
Figure 2 shows an identical longitudinal section at
minimum specific delivery;
Figure 3 shows a cross-section through the pump along
the section line A-A of figure 1;
Figure 4 shows a longitudinal section through the
middle of the shaft and the middle of the
inner rotor along the section line B-B of
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figure 3 at maximum specific delivery;
Figure 5 shows an identical longitudinal section at
minimum specific delivery;
Figure 6 shows a cross-section along the section line
C-C of figure 1;
Figure 7 shows a diagram of the pinion plate with the
axial blading for the axial centrifugal pump,
which blading is fixed on said pinion plate;
Figure 8 shows a diagram of the adjusting member;
Figure 9 shows an alternative embodiment of the
internal gear pump with an alternative
compensating pressure region in a
longitudinal section at maximum specific
delivery;
Figure 10 shows an identical longitudinal section at
minimum specific delivery; and
Figure 11 shows a diagram of the outer rotor of the
alternative embodiment, in the form of a
crown gear.
Since figures 1 to 8 show a common embodiment of the
invention in different views, sections and degrees of__..
detail, figures 1 to 8 are described substantially
together.
Figure 1 shows the variable-volume internal gear pump
in a longitudinal section through the middle of the
shaft and the middle of the inner rotor of the pump at
maximum specific delivery. The internal gear pump has
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a two-part housing which is composed of the actual
housing 7 and a cover 10 of the housing which are
connected to one another by means of screws 19. A
rotor set chamber 40 which has a low pressure chamber
17 with an inlet opening 15 and a high pressure chamber
18 with an outlet opening 16 for a fluid is formed in
the housing 7. The rotor set chamber 40 holds an inner
rotor 2 which is rotatable about an axis of rotation Di
within the rotor set chamber of the housing 7 and can
be driven by a shaft 1 which is passed through the
housing 7 and the cover 10. The inner rotor 2 is
axially moveable on the shaft 1 driving it but is
arranged in a nonrotatable manner only by means of a
feather key 11, substantially runout-free. The rotor
set chamber 40 also holds a rotatable outer rotor 3
having an outer rotor axis of rotation Da arranged
eccentrically relative to the axis of rotation Di, as
shown in figures 3, 4 and 6. The inner rotor 2 has
outer teeth 33 - namely comprising six teeth - and the
outer rotor 3 has inner teeth 34, namely comprising
seven teeth (figures 1 and 3) - such that the outer
rotor 3 rotates with the inner rotor. 2 by means of the
outer-inner teeth 33, 34 in a constant rotational ratio
and, in the case of a rotary drive, forms delivery
cells 30, 31 (figures 1 and 3) in which the fluid is
transported from the low pressure chamber 17 to the
high pressure chamber 18 (figures 1 and 3). The outer
rotor 3 has radial channels 41 arranged in the seven
tooth spaces between the teeth of the inner teeth 34.in
the region of the low pressure chamber 17 and of the
high pressure chamber 18 (figure 2, 3 and 5).
Figures 1, 2 and 5 show an adjusting member 5 according
to the invention which produces an axial movement of
the inner rotor 2. The adjusting member 5 is guided in
an axially moveable manner in the inner teeth 34 of the
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outer rotor 3, the adjusting member 5 having outer
teeth 34a which fit the inner teeth 34 of the outer
rotor 3 with sufficient but little play and are
therefore axially moveable therein while providing a
seal. Figure 8 shows the adjusting member 5 with its
outer teeth 34a in a detailed view. The axial position
of the inner rotor 2 relative to the outer rotor 3 is
adjustable by the axial movement of the adjusting
member 5, with the result that the volume of the
delivery cells 30, 31 is variable.
Distributed uniformly in the inner rotor 2, three
axially acting coil springs 8 are installed between the
shaft 1 driving the inner rotor 2 and the tooth contour
of the outer teeth 33, as shown in figures 1, 3 and 6.
For this purpose, the outer teeth 33 of the inner rotor
2 have a corresponding shape. The three coil springs 8
are supported via a pot-like intermediate member 6
(figures 1 and 5) and a securing ring 12 (figure 1) in
the axial direction on the shaft 1.
The delivery cells 30,- 31 are closed in the axial
direction and in a position opposite to the adjusting
member 5 by a pinion plate 4 and 46, which is evident
in figure 1 and shown in detail in figure 7. The inner
teeth 32 of the pinion plate 4 and 46 fit the outer
teeth 33 of the inner rotor 2 with sufficient but
little play, in such a way that the inner rotor 2 is
axially moveable within the -inner teeth 32 of the
pinion plate 4 and 46. Blading 42 (figures 5 and 7)
corresponding to a centrifugal pump 21 (figure 1) is
arranged on the pinion plate 4 and 46.
For an explanation of the function in the individual
figures, the direction of rotation of the rotor set of
the pump may be in the given direction of the arrows 43
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(figure 3), 44 (figure 6) and 45 (figures 5 and 7), so
that the respective suction side and pressure side
corresponding to the expanding and compressing delivery
cells of the teeth are clearly shown. The intake
connection on the inlet opening 15 which forms the
suction opening is arranged in the cover 10. A suction
space 20 surrounds the pot-like intermediate member 6
and is at the same time the suction side of the blading
42 of the axial centrifugal pump 21. The pressure side
of this axial centrifugal pump 21 is at the same time
the low pressure chamber 17 of the internal gear pump,
which low pressure chamber acts as a suction chamber.
The oil is sucked into the expanding delivery cells 30
against the centrifugal force via the radial channels
41 of the outer rotor 3. In the pump shown in the
drawing, the axial impeller of the centrifugal pump 21
runs in relation to the number of teeth of the rotor
set by a factor of 7:6 faster than the outer rotor 3,
so that the centrifugal pressure in the radial channels
41 of the outer rotor 3 is more than compensated by the
pump pressure of the impeller of the centrifugal pump
21. with increasing speed of the axial centrifugal
pump 21 on the pinion plate 4 and 46 according to
figure 7, the pressure in the suction chamber 17
becomes constantly greater so that vapor and air bubble
formation in the oil and the associated danger of
cavitation are ruled out there, even at the highest
speeds. The same applies to the delivery cells 30 on
the suction side.
The compressing delivery cells 31 (cf. figure 3)
displace the oil into the high pressure chamber 18
towards the outlet opening 16.
In figure 1, the rotor set has the maximum tooth width
when the coil springs 8 are able to push the inner
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rotor 2 and hence the adjusting member 5 completely to
the left, almost up to impact on the housing 7. This
is the case when a very low pressure prevails in the
adjustment space 25, clearly shown in figure 2. It is
expedient if the coil springs 8 are prevented by a snap
ring 13 from pressing the adjusting member 5 axially
onto the housing 7, so that no unnecessary frictional
loss occurs at zero pressure, i.e. when idling. The
distance between this snap ring 13 and the securing
ring 12 in the form of a Seeger circlip ring, both
fixed on the shaft 1, should be chosen so that, between
the packet consisting of, in particular, the adjusting
member 5, the inner rotor 2, the coil springs 8, the
pot-like intermediate member 6 and the pinion 4, 46,
there is still sufficient axial play between the
adjusting member 5 and the housing 7.
If this internal gear pump is now connected on the high
pressure side at the outlet opening 16 to the
lubricating oil circulation, for example of an internal
combustion engine, the oil pressure in the high
pressure chamber 18 increases according to the
absorption curve of the motor with increasing- motor
speed and hence (in the case of a rigid drive) pump
speed. Via channels 23 and 24, which are clearly shown
in figure 2, this high pressure also prevails in the
adjustment space 25 and, depending on the speed level
and hence the high pressure level, pushes the adjusting
member 5 and with it the inner rotor 2 to the right
against the spring force. However, the outer rotor 3
maintains its axial position, owing to the small axial
play between the housing 7 and the pinion plate 4 and
46. The effective tooth width of the rotor set is
reduced thereby and the specific delivery is reduced,
as illustrated in figure 2.
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The hydraulic compressive force in the adjustment space
25 on the adjusting member 5 and hence, via the inner
rotor 2, on the coil springs 8 and the pot-like
intermediate member 6 and also on the axial surface of
the outer rotor 3 via the pinion plate 46, 4 is
supported via the securing ring 12 on the shaft 1. To
prevent the securing ring 12 or the pot-like
intermediate member 6 from running against the cover 10
at the shaft bearing 27 with great force, a
compensating pressure region 22 which is subjected to
high pressure via the channel 23 (figure 2) is provided
on the drive side between a drive wheel 9 (figure 1),
arranged outside the housing 7 on the shaft 1, and the
housing 7. This compensating pressure region 22 is
dimensioned so that the axial force it exerts on the
shaft 1 via a central screw 14 to the left in figure 1
is somewhat smaller than the total hydrostatic axial
force of the rotor system to the right. As a result,
the lubrication and the cooling of an annular sealing
surface 47 between the drive wheel 9 and the housing 7,
which sealing surface seals the compensating pressure
region 22 from the outside, can be optimized.
The alternative embodiment of the internal gear pump,
shown in figures 9 to 11, corresponds in substantial
parts to the embodiments of the internal gear pump
which are illustrated in figures 1 to 8, and it is for
this reason that in some cases only the substantial
differences are discussed below.
The alternative embodiment of the internal gear pump
has the housing 7 with the cover 10 belonging to the
housing. The housing is provided with the rotor set
chamber 40, which has the low pressure chamber 17 with
the inlet opening 15 and the high pressure chamber 18
with the outlet opening 16 for a fluid. The rotor set
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chamber 40 holds the inner rotor 2 which is rotatable
about the axis of rotation D, and can be driven by the
shaft 1. The outer rotor 3a rotatably held in the
rotor set chamber 40 has an outer rotor axis of
rotation arranged eccentrically relative to the axis of
rotation Di. The inner rotor 2 has outer teeth 33, and
the outer rotor 3a has inner teeth 34, such that the
outer rotor 3a rotates with the inner rotor 2 by the
outer-inner teeth 33, 34 in a constant rotational ratio
and, in the case of a rotary drive, forms the delivery
cells 30, 31 in which the fluid is transported from the
low pressure chamber 17 to the high pressure chamber
18. By means of the adjusting member 5 guided in an
axially moveable manner in the inner teeth 34 of the
outer rotor 3, an axial movement of the inner rotor 2
can be produced. The axial position of the inner rotor
2 relative to the outer rotor 3a is variable by the
axial movement of the adjusting member 5 so that the
volume of the delivery cells 30, 31 can be varied.
Here, figure 9 shows the internal gear pump at maximum
specific delivery and figure 10 shows said pump at
minimum specific delivery. The outer teeth 33 of the
inner rotor 2 have a shape such that the axially
effective springs 8 are located between the shaft 1
driving the inner rotor 2 and ' the tooth contour of the
outer teeth 33. Regarding these features, the
alternative embodiment of the internal gear pump of
figures 9 and 10 corresponds to the embodiment of
figures 1 to 8.
The outer rotor 3a likewise has radial channels 41a in
the tooth spaces between its inner teeth 34 in the
region of the low pressure chamber 17 and of the high
pressure chamber 18, but these radial channels 41a are
not formed as radial bores in the middle of the outer
rotor 3a, as can be seen in the first embodiment and as
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can be seen in figure 1, but the radial channels 41a
are present as slot-like radial recesses at the edge of
the outer rotor 3a. In other words, the outer rotor 3a
is in the form of a crown gear, as shown in figure 11
in an individual front view and an individual side
view. An advantage of this arrangement consists in the
easier producability of the radial channels 41a.
in addition, an annular intermediate plate 50 is
provided between the cover 10 and the remaining housing
7, which intermediate plate has passages for the
channels, the pattern of holes corresponding
substantially to the pattern of holes of the housing 7.
The springs 8 are likewise supported via a pot-like
intermediate member 6a and a securing ring 12 in the
axial direction on the shaft 1, but the pot-like
intermediate member 6a has an outer shape - in
particular cylindrical outer shape - such that the
intermediate member 6a touches the cover 10 of the
housing 7 while providing a radial seal. In this
context, a piston ring 48 is provided for sealing.
Thus, a compensating pressure region 22a subjected to
high pressure between the housing 10 and the pot-like
intermediate member 6a and compensating the axial
forces is formed between the pot-like intermediate
member 6a and the cover 10 of the housing 10. In order
to subject the compensating pressure region 22a to high
,pressure, a channel 49 which connects the compensating
pressure region 22a to the high pressure chamber 18 is
provided. This compensating pressure region 22a, too,
is dimensioned so that hydraulic pressure in the
compensating pressure region 22a counteracts the
hydraulic pressure in the adjustment space 25 likewise
subjected to high pressure. In other words, in this
embodiment, the compensating pressure region 22a was
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moved from the side of the drive wheel 9, figure 1, to
the opposite side. This has in particular the
advantage that the drive wheel 9 in the embodiment of
figures 9 to 11 no longer has a sealing function and
can be easily replaced by another drive wheel 9 on the
shaft 1.
