U.S. patent application number 15/920433 was filed with the patent office on 2020-11-12 for power conversion device.
The applicant listed for this patent is Hydracharge LLC. Invention is credited to Jeffrey J. Buschur.
Application Number | 20200355253 15/920433 |
Document ID | / |
Family ID | 1000005178509 |
Filed Date | 2020-11-12 |
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United States Patent
Application |
20200355253 |
Kind Code |
A9 |
Buschur; Jeffrey J. |
November 12, 2020 |
POWER CONVERSION DEVICE
Abstract
A power conversion device in the form of a compressor drive
constitutes a three channel power sharing transmission which allows
power input and/or output from shafts on two of the channels along
with hydraulic, electric or potentially pneumatic power input
and/or output from the third channel. Varying the input and/or
output of hydraulic, electric or pneumatic flow provides a
continuously variable transmission function. Several embodiments of
the power conversion device are described to drive a supercharger
for an internal combustion engine providing a variable ratio
coupling allowing effective use of a centrifugal type compressor
across a broad range of operational engine speeds.
Inventors: |
Buschur; Jeffrey J.;
(Rochester, MI) |
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Applicant: |
Name |
City |
State |
Country |
Type |
Hydracharge LLC |
Auburn Hills |
MI |
US |
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Prior
Publication: |
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Document Identifier |
Publication Date |
|
US 20180202528 A1 |
July 19, 2018 |
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Family ID: |
1000005178509 |
Appl. No.: |
15/920433 |
Filed: |
March 13, 2018 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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14817270 |
Aug 4, 2015 |
9915192 |
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15920433 |
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61999731 |
Aug 4, 2014 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F16H 47/04 20130101;
F02B 39/08 20130101; F16H 61/42 20130101; F16H 61/4035 20130101;
F02B 39/04 20130101 |
International
Class: |
F16H 47/04 20060101
F16H047/04; F16H 61/4035 20060101 F16H061/4035; F16H 61/42 20060101
F16H061/42; F02B 39/08 20060101 F02B039/08; F02B 39/04 20060101
F02B039/04 |
Claims
1. A power conversion device comprising: a first rotating member
disposed for receiving torque, and with rotational motion, power
from a source; a second member disposed to rotate coaxially about
the first member and with said second member comprising a rotating
assembly, wherein said first member and second member are operative
to create a torque relative to each other and with relative angular
motion power; a closed circuit operative to transfer said power to
and from said rotating assembly; and a third member in circuit with
said rotating assembly, said second member coupled to a load to
deliver torque, and with rotational motion power.
2. The device of claim 1, wherein said torque, and with relative
motion between first and second members power, is generated
hydraulically.
3. The device of claim 1, wherein said means to transfer power
comprises a hydraulic circuit.
4. The device of claim 2, wherein the hydraulic means to create
torque, and with relative rotational motion between said first and
second members power, comprises a hydraulic pump consisting of
members affixed to said first and second rotating members.
5. The device of claim 2, wherein the hydraulic means to create
torque and with relative motion between first and second members
power comprises a hydraulic motor consisting of members affixed to
said first and second rotating members.
6. The device of claim 4, wherein the hydraulic pump comprises a
balanced vane design hydraulic pump.
7. The device of claim 3, wherein hydraulic flow is generated when
said first member rotates faster than said second member.
8. The device of claim 3, further comprising an external source of
hydraulic fluid flow in said circuit and flow from said external
hydraulic flow source is transferred to the rotating assembly
enabling the second member to rotate faster than the first
member.
9. The device of claim 7, further comprising a hydraulic motor in
fluid communication to receive said generated hydraulic fluid flow,
wherein said hydraulic motor comprises an output shaft coupled to
add to the output torque transmitted to the load from the second
member.
10. The device of claim 7, further comprising a valve in fluid
communication with the rotating assembly, wherein said valve is
operative to restrict hydraulic fluid flow.
11. The device of claim 9, further comprising a reservoir to store
oil; and a valve to controllably direct said generated flow by i.)
fully blocking it and/or ii.) allowing it to pass through said
hydraulic motor and/or iii.) allowing it to pass directly to said
reservoir and/or iv.) proportioning flow between ii.) and iii).
12. The device of claim 9, further comprising an over running
clutch interfacing said output shaft of said hydraulic motor and
second member by which torque can be transferred only
uni-directionally from the output shaft of said hydraulic motor to
said second member.
13. The device of claim 1, further comprising a speed increasing
mechanism coupled to said second member; and a centrifugal
compressor coupled to said speed increasing mechanism.
14. A power conversion device, comprising a first rotating member
coupled to receive power from a source; a second member operable to
and constrained to rotate coaxially about the first member; a
positive displacement hydraulic mechanism with portions thereof
disposed upon each of the two said members to create a torque
between them and with relative motion a hydraulic flow; a means to
channel the hydraulic flow to and from a rotating assembly
comprised of said first and second members; and a coupling to said
second member to transfer power to a load.
15. The device of claim 14, wherein the hydraulic mechanism
consists of a balanced vane design.
16. The device of claim 14, wherein the hydraulic mechanism
consists of an axial piston design.
17. The device of claim 14, wherein the hydraulic mechanism
consists of a radial piston design.
18. The device of claim 4, wherein the hydraulic pump consists of a
piston design.
19. The device of claim 5, wherein the motor consists of a vane
type design.
20. The device of claim 5, wherein the motor is of gerotor
design.
21. The device of claim 5, wherein the motor consists of a piston
design.
22. The device of claim 14, further comprising: a variable
displacement over center hydraulic pump in fluid communication with
the hydraulic flow moving to and from said rotating assembly,
wherein said variable displacement over center hydraulic pump is
coupled to drive or be driven by said first member.
23. The device of claim 22, further comprising a controlling
mechanism operative to: i.) control displacement of said variable
displacement over center hydraulic pump to produce hydraulic flow
with power drawn from first member and transferring it to the
rotating assembly to overdrive the second member relative to the
first member or, ii.) control displacement of said variable
displacement over center hydraulic pump at or near zero effectively
blocking hydraulic flow from moving from or to said rotating
assembly and substantially preventing rotational movement between
first and second members or, iii.) control displacement in said
variable displacement over center hydraulic pump to receive
hydraulic flow generated by the rotating assembly when the second
member is turning slower than the first member creating a torque to
assist the rotation of the first member.
Description
RELATED PATENT APPLICATION
[0001] This application claims priority to U.S. provisional patent
application Ser. No. 61/999,731 filed 4 Aug. 2014, entitled "Power
Conversion Device" and U.S. utility patent application Ser. No.
14/817,270 filed 4 Aug. 2015, entitled "Power Conversion Device",
soon to issue as U.S. Pat. No. 9,915,192, on 13 Mar. 2018.
BACKGROUND
[0002] The planetary gear set has long been held as the primary
means to provide the function of a power sharing transmission. The
three channels of power input/output are well known to be the ring,
planet carrier and sun. Power can be input/output on each of the
three channels lending itself to the fundamental mechanism used for
automatic transmission functions where locking one of each of the
members to ground allows gear ratio changes as well as to
blend/mingle the internal combustion engine and electric drive
motor(s) for effective use in hybrid vehicle powertrains. In the
case of the planetary set-up used in hybrid powertrains, a varying
drive ratio is produced by having one of the three planetary
channels coupled to a motor/generator which in conjunction with the
controller and batteries can change speed and direction in an
infinitely variable manner. This speed and direction change
provided by the motor generator typically on the highest mechanical
advantage sun gear renders the planetary power sharing transmission
as a CVT. The planetary however unto itself has no ability to vary
the drive ratios to provide CVT function. To date no mechanism has
emerged to allow varying drive ratios without dependency on
frictional interfaces with exception of hydrostatic transmissions
which are essentially a variable displacement pump coupled to the
engine in hydraulic circuit with either a fixed displacement or
variable displacement motor coupled to the drive wheels.
Hydrostatic transmissions have the advantage of needing only
hydraulic hoses to couple the engine to the drive wheels but all
power transmitted is subject to losses incurred in rotary
conversion to pressure and flow and conversely pressure and flow to
torque and rpm at the drive wheels. There would be many potential
uses for an efficient CVT one of which is to enable a centrifugal
compressor to run at near constant speed as engine speed varies to
provide boost for supercharging the engine.
FIELD OF THE INVENTION
[0003] Many internal combustion engines, particularly diesels, are
equipped with a method of boosting the inlet air density in order
achieve greater power per unit engine displacement. Traditionally
there are two main approaches to provide air at a higher pressure
than atmosphere or "Supercharged" to the intake of an engine. The
first is to power an air pump with power drawn from the crankshaft
which has become known in the vernacular as "Supercharging" and the
second is to extract power via a turbine from the exhaust known as
"Turbocharging". Both methods are means of "Supercharging" but the
words Supercharging and Turbocharging have become commonplace to
differentiate the mechanisms in which the compressors are powered.
The vast majority of the diesel engine systems are turbocharged as
power to weight ratios of naturally aspirated diesel engines make
them noncompetitive in the automotive market. Turbochargers are
thus widely utilized in diesel and lately ever increasingly in
spark ignition engines and enjoy a perception in the market that
the power to drive the compressor is essentially free because it is
derived from exhaust engine rather than being drawn from the
crankshaft as in superchargers.
[0004] This is misleading in that the turbines used in the
turbochargers need a significant increase in exhaust back pressure
by which mechanical power can be extracted from the exhaust.
Generally speaking with a typical 65% and 73% adiabatic efficiency
in the turbine and compressor respectively, the required increase
in exhaust backpressure is approximately equal to the level of
intake boost achieved. This essentially causes the engine to
consume more power on the exhaust stroke as the bottom of the
piston is at or near atmosphere and the crank must therefore push
the piston up during the exhaust stroke to move the exhaust out.
For example if a 2.0 liter 4 cycle turbocharged engine is producing
220 ft-lbs of torque but has a 18 psi additional backpressure to
drive the turbo it can be calculated as a Base Mean Effective
Pressure (BMEP) of:
220(ft-lbs).times.24.pi./(2.times.61(in 3/liter)/2(for 4
cycle))=272 psi
[0005] 272 psi thus accordingly a 18/272=0.066 or 6.6% loss
directly attributable to the turbo induced backpressure.
[0006] Further the over pressurized exhaust gas remaining above the
piston when the exhaust valve closes expands and competes with
intake air for volume in the combustion chamber during the intake
stroke. This has a negative impact on engine air throughput and
essentially renders the engine to a slightly reduced displacement
equivalent. The equation below defines this effect.
e.sub.v/e.sub.vb=(k-1)/k+(r-p.sub.e/p.sub.i)/(k(r-1))
[0007] Equation (3) from SAE Paper 730195 Bolt, Bergin and Vesper,
Dept. Mechanical Eng., University of Michigan
[0008] Where:
[0009] e.sub.vb is engine volumetric efficiency when inlet pi and
exhaust pressures p.sub.e are equal
[0010] r is the engine compression ratio
[0011] e.sub.v is the effective efficiency when the ratio is not
unity.
[0012] This effect will reduce engine air mass throughput by some
5% when operating at a typically pressure ratio of 2.0 (or 14.7 psi
boost). It is interesting to note that the lower compression ratios
used by supercharged and turbocharged engines in range of 9 to 10
compared to 11 to 13 for equivalent naturally aspirated engines
exacerbate this effect.
[0013] Overall, despite these little known power losses, turbo
charging is still a more efficient means of driving a compressor
than with power drawn directly from the crank if the turbocharger
can be sized to be optimal at higher ranges of engine speed. If the
turbocharger is optimized for lower end engine speed range these
losses are greatly increased.
[0014] Well known in the industry is the turbochargers delay in
achieving the rotational speed required to develop a meaningful air
pressure rise over atmosphere "or boost" which has become known as
"turbo lag". The delay is caused by the iterative process of time
needed for the increase in engine air flow and thus exhaust
throughput to increase and the inertial resistance of the turbine
and compressor shaft assembly to rotational acceleration.
[0015] Traditionally turbochargers also have had some limitations
as to the boost levels available at lower engine speeds due to
limited exhaust gas velocity and passageway and turbine sizing
necessary to allow maximized power at high engine speeds. Recently
with the renewed emphasis on energy conservation there is much
focus on using turbochargers to allow lower displacement engines
(engine downsizing) in passenger vehicles thus increasing fuel
efficiency. These downsized engine passenger vehicle applications
have driven significant changes in turbo technology to allow the
instantaneous acceleration desirable rather than the sustained high
speed power to which turbochargers are naturally suited.
Turbocharger technology bad been enhanced to overcome these
obstacles in a number of ways. One such trend is use of smaller
exhaust passages and low inertia turbos to allow faster spool up
times to greatly reduce lag but at the cost of torque reducing
restriction of air and exhaust at higher engine speeds. Another
trend in the industry is multiple (aka sequential) turbos utilizing
the ability to employ a restrictive quick spooling turbo for rapid
acceleration and switching to a larger "free breathing turbo" as
time and engine speed allow.
[0016] These downsizing efforts are primarily directed to gasoline
engines where these limitations become more pronounced due to
generally larger operating, speed ranges than in diesel
engines.
[0017] Whether the forced air induction system is powered by an
exhaust turbine or by power directly taken off the crank a
mechanism or compressor must be employed to move air from
atmospheric pressure up into the pressurized or supercharged engine
intake manifold. Compressors can be categorized as being of
positive or non-positive displacement. Positive displacement
devices move a distinct or calculable volume of air from intake to
outlet per stroke or angular input angle. Further within the
positive displacement category are internal and external
compression devices. An internal compression device takes a defined
volume of air from inlet closes off that volume within a chamber
and then reduces the volume of the chamber before discharging the
air at the outlet. An example of this type of compressor is the
Lysholm screw. One disadvantage of internal compression devices is
that even when in bypass mode a considerable amount of work is
exerted on the throughput air raising parasitic losses. An external
compression compressor takes a volume of air from inlet, closes it
off and simply moves it to the outlet port where the volume is then
opened to the pressurized region beyond. As this volume of air
remains at inlet pressure there is an instantaneous backflow into
the chamber as the outlet opens. The volume quickly rises to outlet
pressure and then that volume is pushed into the pressurized region
so the compression takes place "outside" the compressor. This flow
reversal and rapid re-pressurization takes place as every chamber
opens to outlet or several times per revolution tending to create a
pulsing noise. An example of an external type compressor is the
Roots compressor which is highly prevalent in belt driven
supercharging systems. All positive displacement devices loose
efficiency rapidly at higher pressure ratios (>2.0 or boost
>15 psi) as fixed clearances at mechanical interfaces dissipate
more and more of the throughput. Also as this leakage is primarily
a function of pressure, efficiencies of fixed displacement devices
can be quite poor at higher pressures in lower speed ranges where
this leakage is a larger percentage of through-put.
[0018] Non-positive displacement compressors employ either
centrifugal forces or aerodynamics involving airfoils and turning
vanes (aka axial device) as in jet engine compressor sections to
raise air pressure. The volume of air moving through them cannot be
determined strictly from rate of angular displacement but rather a
number of conditions must be known to predict their behavior.
[0019] Centrifugal type compressors are employed exclusively as
coupled with turbines in a turbocharger. Their simplicity of
construction and high efficiencies make them desire-able as air
pumps for these applications. Given that within the moving
interface of the device air velocities are high and pressures are
actually low, leakage is not an issue. In fact most centrifugal
compressors operate more efficiently at pressures higher than
positive displacement devices can reasonably function.
[0020] Centrifugal compressors are also employed on low volume OEM
applications and aftermarket applications in a belt driven
configuration. They create a very potent configuration in upper
engine speed ranges but the inertia it the high speed impellers as
reflected through their step up gearboxes to the drive pulley
create exceedingly high loads on the belt upon rapid engine speed
changes. They are also at the disadvantage that the pressure or
boost created by the spinning impeller is a function of speed
squared. Thus it is a compromise as to the pulley ratio at which
meaningful boost is created in the lower engine speed ranges
with-out consuming huge amounts of power at high engine speeds.
[0021] From the aforementioned it can be deduced that centrifugal
compressors are desirable and that although an efficient means of
powering a forced air induction device turbo charger power is not
without power losses. Lag as mentioned is also a considerable
challenge for an optimized system. Thus it can be understood that a
drive mechanism which can efficiently change the speed of a
centrifugal compressor independent of engine speed would be highly
advantageous.
[0022] Prior Art:
[0023] Several devices have been defined to provide a workable
means for a variable drive for a centrifugal compressor for an
automotive forced air induction system.
[0024] U.S. Pat. No. 8,439,020 to Carlson and Jones defines a
variable ratio drive supercharger assembly which includes a
dry-running CVT coupled to a speed multiplying gear set. This is in
turn is coupled to a centrifugal supercharger. An electric actuator
allows changing of the pulley ratios to allow the centrifugal
compressor to reach operational speeds at lower engine speeds and
reduce ratio to maintain boost pressure capability without
excessive power draw as engine speed increases. Packaging and costs
of the device appear to be prohibitive.
[0025] U.S. Pat. No. 8,366,412 to Grethel defines a means in which
two inputs are combined through use of a planetary and power
sharing transmission as a summer to drive a centrifugal compressor.
One input is coupled directly to the engine via belt and pulley and
the other drive input is from a variable speed hydraulic motor and
pump. By changing the speed of the hydraulic motor in both forward
and reversing directions the output of the planetary and thus speed
of the compressor can be varied. One potential drawback of the
mechanism is the requirement for the planetary gear set to be
rapidly running backward when in the quiescent state potentially
raising parasitic losses and noise levels.
[0026] U.S. Pat. No. 9,080,503 B2 to J. Buschur et al. entitled
"Hydraulic Turbo Actuator Apparatus" describes the use of the
hydraulically driven device in a series configuration with a
minimally restrictive turbocharger is defined which will allow a
very responsive and powerful boosting system to reach boost levels
of 4-5 pressure ratio (PR) to support and enable OEM engine
downsizing trends. An electric supercharger is also considered. A
hydraulic drive assists to increase the acceleration rate of a
turbocharger impeller/turbine shaft assembly and provide a
secondary means of driving the compressor impeller at lower engine
speeds where exhaust gases alone does not generate adequate shaft
speeds to create significant induction boost. The hydraulic circuit
includes a dual displacement motor, which provides high torque for
acceleration yet converts to a single motor for high-speed
operation. When the exhaust driven turbine function allows
compressor speeds, beyond which the hydraulic system can
contribute, a slip clutch allows disengagement of the hydraulic
drive. In an alternative embodiment, the hydraulic drive provides
means of forced induction air alone.
[0027] U.S. Pat. Nos. 8,439,020, 8,366,412 and 9,080,503 B2 are
hereby incorporated herein by reference in their entirety.
SUMMARY OF THE INVENTION
[0028] It is an object of this invention to provide a practical new
means by which power sharing transmission can be constructed
without use of a planetary gear set and additionally a means by
which this power sharing can also offer a CVT function which is
free from dependence on friction yet maintains practical levels of
efficiency.
[0029] It is an object of this invention to provide means to drive
an automotive centrifugal air compressor to supercharge an engine
through a unique rotational summer device such that some of the
power delivered to the compressor is channeled directly from the
engine through the accessory belt or equivalent gear drive and
further some of the power delivered to this same compressor is
borne via means of a hydraulic drive.
[0030] It is a further object of this invention to illustrate means
to provide infinite speed control by adding and subtracting these
two channels of power to the compressor such that engine intake air
pressure can be controllably adjusted independent of and across all
engine speeds.
[0031] It is yet further another object of this invention to
provide definition of certain features making the prior two
objectives feasible, efficient and cost effective.
[0032] These and other features and advantages of this invention
will become apparent upon reading the following specification,
which, along with the drawings, describes preferred and alternative
embodiments of the invention in detail.
BRIEF DESCRIPTION OF THE DRAWINGS
[0033] The present invention will now be described, by way of
example, with reference to the accompanying drawings, in which:
[0034] FIG. 1, is an overall view of a power conversion device
embodying the present invention in the form of a compressor
drive;
[0035] FIG. 2, is a detailed cross section of a portion of the
motor of FIG. 1;
[0036] FIG. 3, is graphical analysis output indicating speed
capability of the drive versus engine speed;
[0037] FIG. 4, is a graphical illustration of drive efficiency;
[0038] FIG. 5, is a hydraulic schematic of the power conversion
device;
[0039] FIG. 6, is a detailed cross section of the fluid coupler of
the power conversion device;
[0040] FIG. 7, is an alternate arrangement of the rotary fluid
couplings of the present invention;
[0041] FIG. 8, is a graphical illustration of comparison of
function of the device versus direct drives;
[0042] FIG. 9, is a schematic illustration of a continuously
variable transmission (CVT) embodiment of the device;
[0043] FIG. 10, is a graphical analysis showing gear ratios and
efficiency as a transmission;
[0044] FIG. 11, is a schematic illustration of an embodiment of the
CVT configuration employing a clutch feature;
[0045] FIG. 12A, is a detailed cross sectional view of a power
conversion device embodying an alternative embodiment of the
present invention in the form of a compressor drive;
[0046] FIG. 12B, is a detailed cross sectional view of the
alternative embodiment of the present invention FIG. 12A;
[0047] FIG. 12C, is a plan view of the alternative embodiment of
the present invention FIG. 12A illustrating the juxtaposition of
the various speed ratio gears;
[0048] FIG. 12D, is a cross sectional view of the alternative
embodiment of the present invention taken on line 12D-12D of FIG.
12C illustrating the details of the gear train and associated
hardware; and
[0049] FIG. 13, is a representative drawing (FIG. 4) of a
representative prior art patent drawing of a viscous damper or
clutch.
[0050] Although the drawings represent several embodiments of the
present invention, the drawings are not necessarily to scale and
certain features may be exaggerated in order to illustrate and
explain the present invention. The exemplification set forth herein
illustrates embodiments of the invention, in several forms, and
such exemplifications are not to be construed as limiting the scope
of the invention in any manner.
DESCRIPTION OF THE PREFERRED AND ALTERNATIVE EMBODIMENTS OF THE
INVENTION
[0051] Referring to FIG. 1, an overall system view shows a Drive
Pulley 1 with shafts 2 and 3 projecting in either direction. Coming
from the pulley 1 to the right is a shaft 3 driving a pump 4 which
feeds flow through a conduit 5 to and from a valve block 6 with a
reservoir 7.1 on a return line 8 and from the other side of the
pulley 1 extends drive shaft 2 affixed to a valve cylinder 7
rotating within a cylinder block 8. The pulley 1 is driven by an
accessory belt (not shown) and turns at a speed roughly 2.times.
that of the associated engine. Rigidly attached to the valve
cylinder 7 are a group of parts clamped together by bolts 9. These
parts which are bolted to the valve cylinder 7 comprise a rotating
motor stack 10 in which is nested a hydraulic motor. An output
shaft 11 of the hydraulic motor turns independently either faster
or slower than the rotating motor stack 10. The output shaft 1 in
turn drives a step up gearing mechanism 12 housed in a gear box 13
which may be typically of 1:12 speed increasing ratio. The output
of the gear box 13 turns a centrifugal compressor shaft 14 which in
turn drives an impeller 15 situated in a compressor housing 16 at a
high rate of speed providing pressurized intake air for an engine
at a compressor outlet 17.
[0052] Referring to FIG. 2, the device will be discussed in greater
detail and the method and significance of function will be
described. The drive pulley 1 again shown on the right is rigidly
attached to a drive shaft 2 which is in turn coupled to the
rotating valve cylinder 7 via an interface 18 which accommodates
some level of misalignment such as the Woodruff keys 19 shown.
Between the pulley 1 and valve cylinder 7 is placed a shaft seal 20
housed in the seal plate 21 which is attached and sealed by a seal
22 in a groove 22.1 against the cylinder block 8. The valve
cylinder 7 has passages 31 and 32 which comprise the supply and
return channels for oil flow in and out of the hydraulic motor
comprised of port plate 25, inner end frame 26, gerotor assembly
27, motor ring 29 and outer end frame 28. Radial grooves 23 and 24
are fly cut on the bore 30 in the cylinder block 8 in which the
valve cylinder 7 turns communicate with the supply and return
channels 31 and 32 in the valve cylinder 7. In this manner the
amount and direction of oil flow to the motor assembly can be
controlled from valves housed within the valve block 6. The inner
end frame 26 and outer end frame 28 house roller bearings 33 and 34
which support the motor shaft 11. The motor ring 29 contains a
larger roller bearing 35 which supports the outer ring 27.1 of the
gerotor 27 (epi-cycloidal gears) which are the motive force of the
motor. A drive pin 36 engages the inner gerotor ring of 27 to the
motor shaft 11. A motor housing 37 rigidly positions and seals the
cylinder block 8 to the bearing block 38 creating a sealed
enclosure 38.1 for the rotating valve cylinder 7 and motor stack
assembly 10. The bearing block 38 further is attached and partially
comprises gear box 13. Passageways through the port plate 25 and
inner end frame 26 allow communication for the oil from the valve
cylinder 7 to the motor ports not shown but formed within inner end
frame 26.
[0053] This mechanism and arrangement embody a device in which
power can be delivered to the motor shaft 11 and thus through
gearbox 13 to the compressor impeller 15 via two means thus
creating a power sharing transmission. The first would be power
turning the motor stack 10 at some torque and speed directly
through the valve cylinder 7, drive shaft 2 and pulley 1 from the
engine. The second would be power defined by that same magnitude of
load torque turning either in an additive or detracting direction
via the hydraulic flow in and out of the motor gerotor 27 turning
it relative to the rotating motor stack 10. It should be understood
by the reader that torque multiplied by angular speed defines
power, thus if the motor gerotor 27 is stationary within the
rotating motor stack 10 the power to the gearbox 13 and compressor
impeller 15 are provided solely by the direct drive from the pulley
1.
[0054] Referring now back to FIG. 1, if hydraulic flow from pump 4
is fed to valve 6 through port 40 and the valve directs flow out of
port 39 through conduit 43 to passage 41 this will drive the flow
through valve cylinder 7 and into motor gerotor 27 in a direction
that is additive to the motion provided directly to the motor stack
10 via transmission of torque from the pulley 1, the two motions
would be additive but the hydraulic drive would only be required to
render that level of power which is required to define the speed
relative to the rotating motor stack 10. It should be noted that if
the construction is such that passage of oil into the motor through
conduits 43 and 41 to fly cut passage 24 (shown on FIG. 2) drive
the motor in additive motion to the direct drive, these passageways
will always be the high pressure side of the motor. The reason is
that the torque load on impeller 15 through gear train 12 will
always create a higher pressure on valve cylinder passage 32 and
fly cut 24 whichever way the gerotor 27 in motor stack 10 is
turning. Therefore conversely if the valve 6 allows flow to leave
the motor opposite the angular motion of the direct drive from the
pulley 1, the gerotor 27 of the motor in the rotating stack 10 will
turn backwards becoming a pump and resistance to that flow in valve
6 from port 39 to port 42 and thus through conduit 8 to the
reservoir 7.1 will define the degree that the motor rotation allows
the input shaft 11 to the gear train 12 to slip relative to the
motor stack 10 detracting from the motion of drive pulley speed 1.
In this manner the gerotor 27 within the rotating motor housing
stack 10 is driving rotating shaft 11 turning relative to the motor
stack 10 which accordingly comprises a mechanical summer as is
commonly accomplished via planetary gear sets today. The three
channels of input/output making the device similar in overall
function to a planetary gear-set are the direct drive shaft 2
coupled to valve cylinder 7, the motor output shaft 11 and the
hydraulic flow moving in or out of motor stack 10 allowing relative
rotary motion between.
[0055] FIG. 3 shows an illustration of the basic function of the
device. The "Speed Motor Stationary" line 44 represents the
impeller speed of the compressor impeller 15 as a friction of
engine speed when the hydraulic motor gerotor 27 mounted in the
rotating motor assembly 10 is stationary. In this illustration the
pulley ratio is at 1:1.6 and speed increaser mechanism has a
1:12.25 ratio. (Note both of these ratios are step-up) The speed of
impeller 15 would thus range from approximately 20000 rpm at 1000
engine rpm to near 130,000 rpm at engine speed of 6500. This would
be problematic as the boost developed by a centrifugal compressor
rises as the square of the speed. This is the basic nature of the
belt driven centrifugal superchargers on the market today and one
of the reasons why they are marginally viable. A single drive ratio
must be chosen that will attempt to provide adequate boost at lower
engine speed while not consuming too much power at the higher
ranges of engine speed. It should be noted the efficiency of this
drive however would be near 100% as only the belt losses and
gearing/bearing loses would be detracting from the power delivered
to the compressor. The "Max Impeller Speed" line 45 depicts the
maximum speed attainable if an 11.3 cc rev pump 4 driven at the
speed of the driven pulley 1 is forcing oil into a 9.8 cc/rev motor
both operating at 87% volumetric efficiency and adding to the speed
of the drive provided directly from the pulley. As can be seen this
line of operation starts at some 40000 rpm at 1000 engine rpm and
exceeds 150,000 rpm in the range of 3750 engine rpm. The "Estimated
Impeller Speed" line 46 represents a function of maximum impeller
speed which may be obtained by adding and subtracting speed
generated by hydraulic motor motion to the speed induced by the
input directly from the pulley drive via control by hydraulic
valves.
[0056] From 1000 engine rpm to 2750 engine rpm at point 47 the
hydraulic motor rotation would add to the rotation induced by the
pulley direct drive with all of the hydraulic flow being produced
by the pump 4 being consumed by the motor gerotor 27 within
rotating motor stack 10. From 2750 rpm to 4500 rpm, valving would
be bypassing flow around the motor otherwise impeller speeds, power
draw and oil pressure would rise unacceptably. At 4500 engine rpm
at point shown 48 the direct drive from the pulley would provide
acceptable input speed to the gearbox and the hydraulics would
essentially be deactivated. This range from 4500 to 6000 engine rpm
is the power band during which the compressor would essentially be
direct pulley gear drive allowing a high level of power
transmission efficiency. Above 6000 rpm at point 49 the valving
would allow backward flow in and out of the motor allowing it to
spin in the opposite direction detracting from the direct pulley
drive speed.
[0057] FIG. 4 indicates the impact on power transmission
efficiencies from the engine crank to the compressor. As previously
stated the motor and pumps in these illustrations have a volumetric
efficiency of 87%. The mechanical efficiencies for the pump and
motor are estimated at 89% and 85% respectively. Overall efficiency
of a hydraulic device is obtained by multiplying volumetric by
mechanical efficiency. The pump is therefore 87%.times.89%=77% and
the motor 87%.times.85%=74%. All these numbers are over
simplifications a both mechanical and volumetric efficiencies vary
with speed, pressure and temperature but for simplicity sake they
are considered constants for these illustrations. Aside from any
loses from bypass flows the hydraulic drive efficiency would then
be 77%.times.74%=57%.
[0058] Now viewing the efficiency line which is scaled on the right
hand axis, it can be seen that the efficiency from 1000 engine rpm
to 3000 rpm is approximately 74% as shown by line segment 50.
Despite the fact that the hydraulic motor is spinning the gearbox
input shaft faster than the direct drive from the pulley the
efficiency is higher than the 57% efficiency of the hydraulic drive
because only a portion of the power is transmitted through the
hydraulics. The remainder moves through the direct drive which is
on order of 99% efficient. In the transition period between 3000
engine rpm and 4500 engine rpm shown at line segment 51 the
efficiency dips to slightly below 50% for a brief range because in
addition to the hydraulic power transmission efficiency of 57%
there are additional losses in the flow bypassed by the valves.
[0059] In the main power band of the engine from 4500 to 6000 rpm
as shown at line segment 52 the power transmission efficiency is
very high on order of 98% because the power is borne by the direct
drive only. A small amount of losses are accounted for in the pump
as it churns unused flow.
[0060] Later at above 6000 engine rpm point 53 the efficiency
trends downward to a low approaching 80% as the hydraulic motor
slipping backward and pumping flow across a control valve
configuration yet to be described represent a loss.
[0061] Referring to FIG. 5, a hydraulic schematic capable of
controlling flow in and out of the motor to facilitate function is
described in FIGS. 3 and 4. Flow from a pump 4 directly coupled to
the engine via pulley and belts or gears provides flow through
conduit 5 to a two position 3 port valve or pump control valve 54.
This valve receives flow from the pump 4 and directs it either into
the motor via conduit 56 or to a bypass 55 back towards the inlet
of the pump. In the relaxed mode the control spool 57 operating in
bore 58 of the valve 54 is in bypass but as the pump solenoid valve
59 is restricted flow is no longer drained from the spring chamber
60 to the right of the spool 57 and it begins to move to the left.
Initially a transition position is reached in which flow is divided
between bypass 61 and outlet 62 feeding the motor allowing the
amount of flow to each to be varied. As the valve spool 57
continues to move to the left eventually all flow is directed to
the motor through valve out 63. Flow from the pump drives the motor
output faster than the direct engine driven rotation as described
related to engine speeds below 4500 rpm as per FIGS. 3 & 4.
[0062] For clarity the rotary fluid couplings 64 and 65 are shown
which allow flow to be channeled in and out of the main rotating
assembly directly driven by the engine.
[0063] A second valve defined as the motor slip valve 66 provides a
controllable means by which the motor comprised of gerotor 27
contained in rotating motor stack assembly 10 can be allowed to be
driven backwards by the reactionary torque of the compressor
gearbox becoming a pump when the compressor is moving slower than
by the direct engine driven rotation. This two position two port
valve controls restriction on a bypass 67 which allows this flow to
move in the opposite direction through the motor and the rotary
fluid couplings 64 and 65. The motor solenoid valve 68 determines
the amount of flow which is allowed to escape from the spring
chamber 69 shown to the left of the motor slip valve spool 70
moving within bore 71. As the spring chamber pressure at 69 is
lowered by this drainage through motor solenoid valve 68 the valve
spool 70 will move to the left allowing a controllable restriction
by which the slip rate of the motor/pump can be controlled.
[0064] There is further shown a relief valve 72 which when
triggered can also drain the spring chamber 69 on the motor slip
valve 66 allowing motor slippage. This mechanism limits the maximum
system hydraulic pressure on the motor/pump and thus the maximum
torque that can be applied to the speed increasing gearbox.
Ultimately this allows a means by which belt and drive loads can be
limited particularly by inertial loads which is a problem on
aftermarket belt driven centrifugal compressors.
[0065] In hydraulic motoring operation as shown in FIG. 3, line 45
of the motor slip valve 66 would be closed while the pump control
valve 54 is forcing flow into the motor. Conversely, the pump
control valve 54 would be in bypass or relaxed mode when either the
motor slip valve 66 is closed causing a direct locked drive from
engine to compressor or when the motor slip valve 66 is open
allowing the compressor to freely slip back. In this mode there may
also be advantage to decoupling pump 4 from the engine drive
reducing parasitic losses.
[0066] Further there would be provided a restriction 73 in the
return line 74 which is common to the pump and the motor which
would provide a pressure drop for controlling the flow rate through
a liquid to liquid cooler. As flows and power levels in the device
increase the delta pressure across the restriction would rise and
more flow would be directed to the glycol coolant circuit 75. Thus
at lower flow rates and accordingly lower engine speeds flow to the
glycol coolant circuit 75 would be reduced allowing oil to stay
warmer reducing viscous losses. The restriction 73 would also
provide an atmospheric pressure clamp (reference point) to keep the
system from changing pressures due to volumetric
expansion/contraction of the fluid due to pressure and temperature
changes. Small amounts of oil would come in and out of the system
as make-up from the reservoir 76. As the flow slowed through the
diverging passage 77 leaving the restriction velocity would be
traded for pressure via Bernoulli Effect, providing an oil inlet
boost mechanism to prevent cavitation in the pump and/or the
motor/pump when in pump mode.
[0067] FIG. 6 illustrates a slightly enlarged view of a section of
FIG. 2. Illustrated is the valve cylinder 7 which rotates inside
the cylinder block 8. Passageways 31 and 32 in the valve cylinder 7
are in communication with passages and valves in the cylinder block
8 via two fly-cut grooves 23 and 24 circumventing the bore 30 in
the cylinder block 8 in which the valve cylinder 7 rotates. It
should be noted and is of great design convenience that despite
bidirectional flow in and out of the valve cylinder 7 that one set
of passages in this configuration 32 is always high pressure and
the other 31 is always return or low pressure as whether the motor
is being driven by the pump or slipping back in rotation the
direction of delta pressure across the motor never changes.
[0068] There is shown on FIG. 6 dimensions A 78 and B 79 as well as
valve cylinder diameter 80 (VCD). Dimensions A 78 and B 79
represent the linear distance of sealing land on either side 81 and
82 of the high pressure flow groove 24 fly cut within cylinder
block 8. In the design as illustrated either end of the valve
cylinder bore 30 are drained to tank and represent a power loss and
thus should be minimized but the leakage also serves as forced
fluid feed to allow a hydrodynamic rotating interface of the
cylinder block 8 to the valve cylinder 7. The following well known
equation defines leakage between a cylinder and a bore in which it
is centered.
Leakage=((.pi.DC.sub.r.sup.3P)/uL)k
[0069] Where: D is the diameter of the bore which in this case is
VCD,
[0070] Cr is the radial clearance,
[0071] P is the pressure differential,
[0072] u is viscosity in centipoise,
[0073] L is the length of in this case A and B, and
[0074] k is coefficient of convenience for unit's conversion,
[0075] It can be understood from the equation that the radial gap
is the most significant variable in the function and that is
subject to manufacturing variance. It is also obvious that D or VCD
in this case should be minimized however again looking at FIG. 6 it
should be realized that the cross section defined by VCD needs to
be large enough to allow two passages of sufficient cross section
to channel flow to and from the motor. If VCD is enlarged there is
an increase in the power loss required to turn it within the valve
block due to the shearing of the oil at the interface. The commonly
known equation for calculating this loss is called the Petroff's
equation and is as follows:
Power Loss=(.pi..sup.3/C.sub.r)D.sup.3N.sup.2uLk
[0076] Where: N is the relative speed at the rotating interface,
and
[0077] All other variables are the same as in the above leakage
equation.
[0078] Here the conflict between wanting a small C.sub.r to reduce
leakage loss and a large C.sub.r to reduce shear losses are shown
to be directly at odds with each other. Also L wants be large to
reduce leakage but also small to reduce shear losses. It has been
found that for a system with 10-25 gpm (gallons per minute) in flow
and 10-30 hp (horse power) capacity a reasonably workable
combination can be achieved with D=32 mm, Cr=0.030 mm and L=10
mm.
[0079] Referring to FIG. 7, an alternate embodiment in which D or
VCD can be drastically reduced by making the two passageways in 83
and out 84 of the motor gerotor 27 come from either end rather than
both from one side. This requires that on the motor shaft output
end 85 the flow needs to make two transitions 86 and 87 from the
central shaft hole to an interfacing shaft bores as shown. Despite
the increased leakage this double rotary coupling incurs the much
smaller D more than compensates. With this configuration D can
easily be reduced to the range of 19 mm while still leaving more
than adequate internal cross section to allow large flow
volumes.
[0080] Referring to FIG. 8, the functionality of the subject
invention is illustrated in comparison to a direct belt driven
centrifugal compressor. The line 88 represents the desired maximum
speed capability of the ideal compressor system as ideal boost
capability would remain the same independent of the engine as it
closely correlates to the engine torque. It remains essentially
constant across, engine speed with exception that there is a slight
rise as engine speed increases to represent the additional internal
restriction losses as the air mass flow increases. If a centrifugal
compressor is driven directly by the belt its speed will be
directly proportional to engine speed as shown by line 89 thus will
be far from the ideal speed. Despite this deficiency the power
transfer efficiency to the compressor from the engine is high as
the belt and pulleys constitute minimal loss. The middle region 90
around the intersection of lines 88 and 89 represents a region of
high power transmission efficiency because the power borne by the
hydraulics is minimal. The two triangles of "Speed Added by
Hydraulic Motor" 91 and "Motor Slippage Controlled Hydraulically"
92 represent power transfers subject to hydraulic losses in the
pump and motor.
[0081] Now looking at this chart (FIG. 8) in a different light it
can be recognized that the subject device is essentially a
continuously variable transmission. Assuming engine speed or power
input traverses from say 700 rpm to 7000 rpm the ability to
maintain a near constant output to the compressor means the subject
device is ranging through an input/output ratio of approximately
10.times.. In the midrange of engine speed the ratio can be equated
to approximately 5.times. with minimal losses as it is essentially
a direct drive mechanism. At the highest of engine speed the ratio
has been reduced to 1.times.. At near off idle the speed ratio is
10.times..
[0082] The device as discussed to this point is configured to
simply let the motor slip backwards becoming a pump when the direct
drive speed is higher than the desired speed. The power transfer as
shown in the upper triangle is simply dissipated across the delta
pressure of the metering motor slip valve. The flow could be
channeled back through the pump which would then become a motor
allowing torque generated in the direction of lowering the power
input to the direct drive.
[0083] FIG. 9 shows a simplified circuit 93 in which a variable
displacement over center pump/motor 94 is substituted for the fixed
displacement pump 4 used in the previous embodiments. In this
manner the device becomes a high efficiency continuously variable
transmission (CVT). The pump control valve and the motor slip valve
are removed. The motor/pump 95 "Device A" however is still mounted
with its reference frame being the rotating assembly driven
directly by the engine. The over center pump motor 94 has an output
shaft 103 which is mechanically coupled with an input shaft 104 of
motor/pump 95 as indicated by dotted line 105. Rotary fluid
couplings 64 and 65 are still employed to carry flow in and out of
this rotating motor/pump assembly 95. In operation, the variable
displacement puny/motor 94 "Device B" would still provide flow when
the output of the CVT is desired to be above the output speed of
the rotary input driven directly by the motor. During this mode of
operation power provided directly by the engine and hydraulically
delivered power are summed to provide an output. Displacement of
device B 94 would be varied to allow adjustment of drive ratio.
Efficiency would remain high as a large percentage of the power
would be transferred at near 100% efficiency. During mid-drive
ratio range the pump (Device B) would go to neutral position thus
stopping all flow and the motor (Device A) 95 on the rotating
assembly would lock and turn as the same speed as the rotation
driven directly by the engine. During this mode of operation the
power transfer efficiency would be very high as the assembly would
be rendered essentially a locked shaft. When speed driven by the
engine is higher than desired output the motor (Device A) on the
rotating group would be allowed to slip backwards creating a
pumping action. The variable displacement pump/motor (Device B)
would now go over center becoming a variable displacement motor.
The displacement would be adjusted such that the flow being
produced by the slippage of the motor (Device A) on the rotating
group would allow the pump (Device B) in it's now motoring mode to
match the speed by which it was being driven by the engine shaft
96. In this manner the torque load applied by the engine to the
direct drive would be reduced. It should be noted that an electric
or pneumatic motor could gain a similar effect if mounted as
described with its reference frame rotating except power (torque)
density would be much lower than with a hydraulic device.
[0084] Equations defining such a CVT would be as follows:
Gear Ratio=N.sub.In/N.sub.Out=1/(1+D.sub.B/D.sub.A), and
Hydraulic Pressure=(24.pi.T.sub.Out)/D.sub.A.
[0085] Where: T.sub.Out=Transmission output Torque (ft-lbs),
[0086] Da=Displacement of Device A (cubic inches per rev)
[0087] Db=Displacement of Device B (cubic inches per rev)
[0088] Note--Db is defined as a negative value when Device A is
slipping relative to the input shaft and a positive Db is
indicative of the Device A output being faster or additive to
rotation of the input shaft.
[0089] Hydraulic Pressure is in units of psi.
[0090] Efficiency (Device A Output Rotating Faster than Input
Shaft).
Efficiency = D a .times. E ma .times. ( 1 + D b .times. E vb D a
.times. E va ) ( D a .times. E ma + D b E mb ) ##EQU00001##
[0091] Efficiency (Device A Output Rotation Slower (Slipping) than
Input Shaft)
Efficiency = D a E ma .times. ( D a .times. E va - D b E vb ) D a
.times. E va .times. ( D a E ma - D b .times. E mb )
##EQU00002##
[0092] Where: [0093] E.sub.va=Volumetric (slip) Efficiency of
Device A [0094] E.sub.vb=Volumetric (slip) Efficiency of Device B
[0095] E.sub.ma=Mechanical (torque) Efficiency of Device A [0096]
E.sub.mb=Mechanical (torque) Efficiency of Device B
[0097] Typically automotive transmissions start in first gear in
the range of a 4:1 ratio and transition to 1:1 locked in normal
drive mode. Overdrive is when the ratio goes below 1:1 to perhaps
0.7:1 indicating the output of the transmission device is rotating
faster than the input.
[0098] The graph in FIG. 10 illustrates a CVT function for an
automobile application that can be accomplished with such a
described device. It is similar or somewhat better in efficiency to
what is accomplished with discrete planetary gearing in
transmissions used by automobiles today. Upon acceleration the CVT
would start with the displacement of Device B less than zero shown
at segment 97 indicating that it is in pumping mode allowing Device
A to slip backward. As the vehicle accelerates the displacement of
Device B would be reducing approaching zero from the negative side
at point 98 where it would remain for the normal drive function. It
should be noted that this would provide a locked 1:1 ratio or
straight through as is called yielding high efficiency point 99 at
normal driving conditions. Under certain conditions the
displacement of Device B could be shifted to positive segment 100
to drive the Device A faster than the direct drive engine input
yielding what is defined as overdrive. In this mode the efficiency
segment 101 would be quite high still as only a fraction of the
power transmission would be subject to losses in Devices A and B.
Thus is provided an efficient CVT function which is in some ways
similar to hydrostatic transmissions but yielding much higher
efficiency particularly in locked drive mode as power can be
transmitted directly without being borne by the hydraulic fluid
motion.
[0099] FIG. 10 shows use of a Device A of displacement 0.5 cubic
inches per rev and Device B have a maximum displacement of 0.325
cubic inches per rev but these displacements could be scaled up and
down for varying CVT torque capacities and hydraulic fluid pressure
levels in the preferred ratio of Device B being approximately
50-80% that of Device A.
[0100] It should also be noted that if an additional valve 102 is
added as per FIG. 11 allowing the motor on the rotating group
(Device A) to slip it can provide a clutching function which is
easily accomplished with electronic control as per the motor slip
valve described in FIG. 5. This clutching function could be very
smooth, easily electronically controlled and virtually free of wear
of any kind.
[0101] FIGS. 12A-12D illustrate another embodiment of the
invention. Input shaft 106 is driven by gear 107 through drive
wheel 108. A key 109 locks wheel 108 to shaft 106. Shaft 106 is
supported by needle bearing 110 near the drive gear 107 and by
another roller bearing 112 housed in low pressure valve plate 111.
Shaft 106 is engaged to a vane pump rotor 113 through a series of
splines 114. Rotor 113, a cam ring 115, low pressure valve plate
111 and a high pressure valve plate 116 comprise a balanced vane
pump assembly stack as is well known in the art of hydraulic
design. Ten vanes which follow the surface of cam ring 115 moving
radially in and out in slots as shown 117. Normally this type pump
assembly stack would be contained in a stationary housing and would
pump flow proportional to the speed of shaft 106. In this
embodiment, the cam ring 115, high and low pressure plates 116 and
111, respectively, along with other components as explained later
all constitute a rotating assembly which is free to rotate
coaxially with shaft 106.
[0102] It should be noted that although a balanced vane pump is
described and illustrated within this application, an axial piston
hydraulic device would work as well and likely would be more
efficient.
[0103] As the rotor 113 turns with its vanes moving accordingly
along the surface of cam 115, fluid would be pulled in at ports
118a and 118b and expelled at ports 119a and 119b. Again, in
contrast to normal construction, this pumping assembly is housed as
clamped between low pressure end cap 120 and high pressure end cap
121 which along with several other components are free to turn as a
rotating assembly coaxially with shaft 106. At the interface of end
caps 120 and 121 is clamped and anchored a gear 122 which
circumvents the pumping assembly. Several bolts 123 hold this
assembly together. This rotating assembly further includes a high
pressure seal sleeve 124. It contains seal grooves 125 and 126
which along with seal groove 127 in high pressure valve plate 116
contain and direct pump output flow from the pumping action of the
vanes out to radial holes 128 in high pressure end cap 121.
[0104] High pressure end cap 121 and a manifold 130 comprise two
journal and bearing interfaces, the outer being 129a and the inner
being 129b. The journal and bearing surfaces 129a and 129b are
formed at a bi-section in a bore in manifold 130 due to a fly cut
131. This fly cut 131 along with holes 119a (?) in high pressure
end cap 121 and the bearing journals between manifold 130 and cap
108 direct high pressure pump flow from the region between the high
pressure seal sleeve 124, high pressure end cap 121 and high
pressure valve plate 116. Thus these components comprise the high
pressure rotary coupling.
[0105] The entire aforementioned clamped and bolted rotating
assembly is not, stationary but is free to rotate coaxially with
shaft 106 as it positioned by a ball bearing assembly 132 supported
and positioned in a cover 154 and journal bearings 129a and 129b.
As there is no polymeric seal on the high pressure rotary coupling
there will be some significant leakage flow out past journal 129b
into the area between manifold 130 and high pressure end cap 121.
Oil leaking past inner journal 129b will drain down into reservoir
volume 133. Oil leaking past outer journal 129a would be contained
in a gear cover 177 (FIGS. 12C and 12D) and eventually channeled
down to the reservoir volume 133.
[0106] It should be noted that the input torque at shaft 106 is
essentially balanced, aside from frictional losses and inertial
effects, by the output torque created by the tangential force at
the engagement of gear 122 and a gear 158. This tangential force
times the moment arm of the distance from the center of rotation
being that of shaft 106 define a moment or torque which equals the
input torque on shaft 106. Thus, these torques are basically
equivalent but the speed of rotation of shaft 106 and gear 122 on
the overall rotating assembly need not be and represent a slippage
rate which determines generally the amount of flow moving through
the pump.
[0107] FIG. 13 is taken from a prior art U.S. patent (U.S. Pat. No.
1,128,447) illustrating a representative viscous damper or clutch.
This is a good reference for comparison and contrast to the art
described in this application. The two halves of this assembly are
coaxial and further comprise a group of concentric interfaces in
which a viscous fluid such as oil or silicon fluid creates a slip
due to sheer of the fluid. These types of devices are common as
rotational dampers and torque limiters.
[0108] In FIG. 13, the prior art reference numerals correspond
to:
[0109] 6'--body
[0110] 7'--detachable head
[0111] 11'--shaft
[0112] 12'--disk
[0113] 13'--annuli
[0114] 14'--disk
[0115] 15'--annuli
[0116] 19'--holes
[0117] Note: apostrophes ['] are added to the FIG. 13 reference
numerals to distinguish similar reference numerals describing the
present invention
[0118] A variation on this arrangement in which the amount of fluid
contained in the grooves is varied to control the slip torque is
used for automotive engine driven radiator fan clutches. The
torques on the input and output shafts in this case are generally
equal also, aside from dynamic effects, but the slip rate times
this torque creates power that generates heat within the fluid
being sheared which aside from convection to surrounding air has
nowhere to go thus the accumulated energy results in rapid
temperature rise.
[0119] Another known device with similar singular input and output
shaft torque interfaces are torque converters. They work on
primarily fluid momentum instead of viscous shear so their torque
vs speed transfer functions are different but the slip between
input and output halves causes heat just as in the generic viscous
couplings. Often torque converters are equipped with a means to
send the fluid out of the rotating assembly to be cooled but this
is not a channel through which productive power is moved.
[0120] Directing back to the embodiment illustrated in FIGS.
12A-12D, oil flow generated by the pumping action caused by the
slip rate between shaft 106, gear 107, drive plate 108 and pump
rotor 113 and the opposing half of the hydraulic assembly now
recognized as the rotating assembly (comprising of parts 111, 115,
116, 124, 121,120,122 and 123) is not churned into heat as in
viscous clutches or torque converters but is rather channeled off
the rotating assembly for useful purposes.
[0121] Now best seen in FIG. 12A oil moving into fly-cut 131 in
manifold 130 is channeled into a bore 134 which has a slightly
smaller diameter than spool bore 135. Within spool bore 135 is an
axially moving spool shown in forward most position by 136a and
rearmost position by 136b. A spring (not shown) is contained in the
rear of spool bore 135 and biases spool 136 leftward. High pressure
oil from fly-cut 131 is channeled through passage 137 and
restriction orifice 138 on its way to the volume behind the spool
139. An electrically controlled valve 140 controls the restriction
from volume 139 to drain passage 141. This comprises a pressure
balance on spool 136 in which the pressure in the volume of bore
134 is controllably fixed by the control signal to valve 140. This
valve arrangement is well known in the hydraulics industry as a
dual stage pressure control valve.
[0122] If the pressure control valve 140 is relaxed allowing flow
out of volume 139 in excess of the flow moving in through orifice
138, the spool 136 will move to the right towards spool position
136b. As the spool 136 retracts to the right oil can first pass
into hydraulic gerotor motor 145 through inlet port 142. Hydraulic
motor 145 rotates about its axis at shaft bore 143 within manifold
130. Oil discharged from motor 145 at outlet port 146 is directed
down passage 144, As this passage moves down below oil level shown
at 147 it turns and goes into a convergence region 148. This
accelerates the oil to a velocity higher than passage 144 reducing
its pressure according to the Bernoulli principle. The oil moves
through a constant cross section 149 for a length in which are
several smaller holes 150. These holes 150 present a low pressure
zone to oil in the reservoir below oil level 147 and generally all
leakage oil is pulled back into action. The oil then encounters a
divergence at 151 of approximately 6 degrees included angle. The
oil slows and pressure rises again according to Bernoulli principle
in this region.
[0123] If the spool 136 retracts to position 136b, an opening is
presented to the oil to move into passage 144 directly without
passing through motor 145. In either case, as the oil leaves
divergence section 151 it enters passage 152 which leads through a
hole 156 to an end cavity 153 comprised of a pocket in cover a 154
and a closure with low pressure end plug 155. Oil is then pushed by
pressure developed by region 151 and sucked by pumping action
through an inlet in the end of low pressure end cap 120. A
polymeric seal 157 is positioned between cover 154 and low pressure
end cap 120 ensuring no oil leakage or air entrainment through this
second low pressure rotary coupling.
[0124] If valve 140 remains in relaxed position allowing oil to
depart volume 139 faster than it enters via orifice 138 then the
spool 136 will move to position 136b allowing oil to return
directly to the reservoir 133 through passage 144 and oil produced
by the pumping group by the rotation of shaft 106 will not generate
pressure at ports 119. If there is little or no pressure build the
torque reaction on cam ring 115 will be minimal and the assembly
clamped between end cap parts 120 and 121 will turn slowly within
its bearings if at all. Gear 122 clamped onto the assembly will
provide little if any drive power to its meshing gear 158, which is
mounted on the output shaft of gerotor motor 145 centered at bore
143 via an over running slip clutch visible and later explained
further in FIGS. 12C and 12D.
[0125] Referring now to FIGS. 12C and 12D, the gear train and
associated hardware, which in this embodiment is configured to
propel a centrifugal compressor for automotive forced air
induction, will be described. Gear 122 mounted circumferentially
around the aforementioned rotating assembly is in gear mesh with
gear 158. This mesh can be clearly seen in FIG. 12C at point 159,
but gear 122 is obscured from view in FIG. 12D. Gear 158 is coupled
to the output shaft 160 of gerotor motor 145 but is only partially
shown in FIG. 12D. At the interface of gear 158 and shaft 160 is an
over-riding roller clutch 161. The clutch 161 allows the shaft 160
to drive the gear 158 but slips if the gear is over speeding the
shaft 160. Gear 158 is also in mesh with a smaller gear 162 which
is mounted on an intermediate shaft 163 which is supported by
bearings 164a and 164b. Permanently fixed to shaft 163 is also a
gear 165 which in turn is in mesh with a smaller gear 166 mounted
on a compressor shaft 169. The compressor shaft 169 is supported by
ball bearings 167 and 168. Finally permanently attached to shaft
169 is compressor impeller 170. Thus gear 122 drives gear 158 at
approximately a 1:1 ratio. Gear 158 drives gear 162 at a step up
ratio of approximately 1:3.5. Further gear 165 drives the
compressor shaft 169 and in turn impeller 170 at a step up ration
of an additional 1:3.5 ratio to intermediate shaft 163. Overall the
speed up gear ratio is thus some 1:12 from gear 122 to impeller
170.
[0126] On the other end of the powertrain in one instance a belt
(not shown) is driven by a serpentine pulley 171 attached to a
drive shaft 172 which is supported on one end by a main bearing 173
and on the other end by another bearing obscured from view in FIG.
12D. Attached to shaft 172 is a larger gear 174 via a bowl shaped
flange 175. Gear 174 is in mesh with and drives gear 107 shown in
FIG. 12C in dotted line as it is behind gear 122. A drive shaft
seal 176 is mounted within and sealing in turn against gear cover
177. The gear cover 177, in turn, is mounted against and seals
against manifold 130 shown in FIG. 12B. In this manner the entire
gear train and general mechanism is in a sealed enclosure and the
leakage through bearing journals 129a and 129b is entirely
contained.
[0127] In operation, when valve 140 is restricted and a load is
meshed with gear 158, such as the gear train coupled to a
centrifugal supercharger impeller as described, pressure will build
at pump exit ports 119a and 119b. Reaction torque on cam ring 115
is directly proportional to this pressure and, thus, also
represents the available drive torque at gear 122. A slip rate
between the rotational speed of the input shaft 106 and that of the
rotating assembly turning in bearings 132 and 129a will define the
flow produced by the vanes in slots 117 of rotor 113 as they follow
cam ring 115. This flow will be metered by spool 136 either through
the motor 145 or directly into passage 144 at a pressure drop
generally equal to pressure at ports 119 relative to the pressure
of reservoir volume 133. If the flow is adequate to propel the
gerotor of motor 145 to the speed of gear 158 the over-riding
clutch 161 (see FIGS. 12C and 12D) will engage providing a
productive torque which will increase the overall efficiency of the
mechanism. When flow is lower than this volumetric rate, the
pressure drop times the flow will represent a power loss. If the
flow is above the volumetric flow required to propel the motor at
the same speed as gear 158, the excess flow times the pressure drop
across the spool metering interface to passage 144 will represent a
loss. Overall, however, the efficiency of power transfer from input
shaft 106 to engaging gear mesh on gear 158 will be reasonably high
as only the slip rate represents a substantial loss and a large
portion of that slip power will be recovered by motor 145. This
provides an efficient controllable continuously variable drive
mechanism.
[0128] The present invention is intended for application in varied
automotive vehicle applications and will be described in that
context. It is to be understood, however, that the present
invention could also be successfully applied in many other
applications. Accordingly, the claims herein should not be deemed
limited to the specifics of the preferred embodiments of the
invention describer hereunder.
[0129] The following documents are deemed to provide a fuller
disclosure of the inventions described herein and the manner of
making and using same. Accordingly, each of the below-listed
documents is hereby incorporated in the specification hereof by
reference in their entirety.
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[0132] U.S. Pat. No. 5,535,845 to J. Buschur entitled "Automotive
Hydraulic System and Method".
[0133] U.S. Pat. No. 5,561,978 to J. Buschur entitled "Hydraulic
Motor System".
[0134] U.S. Pat. No. 5,669,461 to J. Buschur entitled "Automotive
Hydraulic System and Method".
[0135] U.S. Pat. No. 5,687,568 to J. Buschur entitled "Hydraulic
Motor System".
[0136] U.S. Pat. No. 5,778,693 to M. Mientus entitled "Automotive
Hydraulic Engine Cooling System with Thermostatic Control by
Hydraulic Actuation".
[0137] U.S. Pat. No. 5,881,630 to J. Buschur et al. entitled
"Apparatus and Method of Controlling Fluid Flow between a Plurality
of Vehicle Components".
[0138] U.S. Pat. No. 5,946,911 to J. Buschur et al. entitled "Fluid
Control System for Powering Vehicle Accessories".
[0139] U.S. Pat. No. 5,960,628 to K. Machesney et al. entitled
"Hydraulically Powered Fan and Power Steering in Vehicle".
[0140] U.S. Pat. No. 5,960,748 to J. G. Lewis entitled "Vehicle.
Hydraulic Component Support and Cooling System".
[0141] U.S. Pat. No. 6,016,657 to J. Buschur entitled "Automotive
Hydraulic System and Method".
[0142] U.S. Pat. No. 6,021,641 to J. Buschur et al. entitled
"Hydraulically Powered Fan System for Vehicles".
[0143] U.S. Pat. No. 6,158,216 to 3. Buschur et al. entitled
"Hydraulically Powered Fan System for Vehicles".
[0144] U.S. Pat. No. 6,308,665 B1 to J. G. Lewis entitled "Vehicle
Hydraulic Component Support and Cooling System".
[0145] U.S. Pat. No. 6,612,822 B2 to J. Busch are et al. entitled
"Hydraulic Motor System".
[0146] U.S. Pat. No. 6,629,411 B2 to J. Buschur et al. entitled
"Dual Displacement Motor Control".
[0147] U.S. Pat. No. 7,608,011 B2 to Grabowski et al. entitled
"Hydrogen Fueled Hybrid Powertrain and Vehicle".
[0148] U.S. Pat. No. 7,490,594 B2 to Van Dyne et al. entitled
"Super-Turbocharger".
[0149] U.S. Pat. No. 7481,056 B2 to Blaylock et al. entitled
"Turbocharger with Adjustable Throat".
[0150] U.S. Pat. No. 7,111,704 B2 to Johnson entitled "Hydrostatic
Drive Apparatus for a Road Vehicle".
[0151] U.S. Pat. No. 6,502,398 B2 to Kapich entitled "Exhaust Power
Recovery System".
[0152] U.S. Pat. No. 6,412,278 B1 to Matthews entitled
"Hydraulically Powered Exhaust Gas Recirculation System".
[0153] U.S. Pat. No. 5,724,949 to C. Liang entitled "Hydraulic
Drive for a Pressure Wave Supercharger Utilized with an Internal
Combustion Engine".
[0154] U.S. Pat. No. 5,346,364 to D. Kapich entitled "Very High
Speed Hydraulic Turbine Drive". And
[0155] U.S. Pat. No. 4,729,225 to J. Bucher entitled "Turbo-Charged
Internal Combustion Engine with Exhaust Gas Energy
Recuperation".
[0156] U.S. Pat. No. 9,080,503 B2 to J. Buschur et al. entitled
"Hydraulic Turbo Actuator Apparatus".
[0157] It is to be understood that the invention has been described
with reference to specific embodiments and variations to provide
the features and advantages previously described and that the
embodiments are susceptible of modification as will be apparent to
those skilled in the art.
[0158] Furthermore, it is contemplated that many alternative,
common inexpensive materials can be employed to construct the basis
constituent components. Accordingly, the forgoing is not to be
construed in a limiting sense.
[0159] The invention has been described in an illustrative manner,
and it is to be understood that the terminology, which has been
used is intended to be in the nature of words of description rather
than of limitation.
[0160] Obviously, many modifications and variations of the present
invention are possible in light of the above teachings. It is,
therefore, to be understood that within the scope of the appended
claims, wherein reference numerals are merely for illustrative
purposes and convenience and are not in any way limiting, the
invention, which is defined by the following claims as interpreted
according to the principles of patent law, including the Doctrine
of Equivalents, may be practiced otherwise than is specifically
described.
* * * * *