Note : Les descriptions sont présentées dans la langue officielle dans laquelle elles ont été soumises.
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Field of the Invention
This invention relates to an engine-hydromechanical trans-
mission unit providing a torque converter in combination with aplurality of mechanical gear steps and, more particularly,
to a method for providing braking in such a transmission
which can be regulated, according to demand, within pre-
determined limits.
Background of the Invention
Hydromechanical transmissions have been utilized
to obtain braking torque as desired. However, systems used
up to now, such as, for example, that disclosed in commonly
assigned U.S. Patent No. 3,261,232, do not provide any
capability for regulating braking except hy connecting
different mechanical gears. Conventional
hydromechanical transmissions have heretofore utilized a
separate "retarder". Although such arrangements provide
regulation or modulation of the braking torque, the retarder,
which comprises a variably filled hydrodynamic coupling, pro-
vides considerable resistance against rotation when not in
use and only filled with air. While there are techniques
for reducing this loss of the air filled retarder, the arrange-
ment for accomplishing this requires considerable space and
is heavy and expensive.
As stated above, the prior art system e~n~-e~-e~ie~-
L~
would not regulate braking itself. At best, one could obtain
different degrees of braking capacity by different settings
of the torque converter and by utilizing different mechanical
gear connections~ However, the method of the present invention
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provides for regulatingthe braking force betweenwide limitsand with
less shock when changing between mechanical gear ratios, and
does not depend on any additionalmechanical structuresexcept that
concerning the automatic control for providing a predetermined
braking setting.
Summary of the Invention
As is evident from the foregoing, a basic purpose of
the present invention is to extend the ~raking capacity of an
engine-transmission unitwithout addingany additional structure
in order to obtain the braking force. At the same time, the
invention makes it possible to vary the braking torque within
wide limits. It is noted that these limits are greater than
those provided with separate hydraulic retarder units such as
mentioned above and that the invention also avoids the losses
associated with such retarder units when not connected. In
addition, the invention provides softer shifts between the
different braking fields defined by the gear ratio of the
gear which is connected. A further purpose of the invention is
to provide automatic control of the braking torque in such a
way that the driver can manually set the desired braking
capacity independently of speed and according to a predetermined
relationship between vehicle speed and braking capacity.
In general, the invention is concerned with a method
used with an engine transmission unit for a vehicle wherein the
transmission comprises a torque converter having a direct drive
~lutch, a guide vane member which can be released or fixed
against rotation (or even be connected to be driven backwards)
and a mechanical gear transmission of the power shift type in
series with the torque converter. According to the invention,
the braking torque is obtained by running the turbine member
of the torque converter at a higher speed than the primary side
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;
of the torque converter (the pump member) which is rotationally
connected to the engine, i.e., by overspeeding the turbine
member. This overspeeding of the turbine member is carried
out wit:h the direct drive clutch released and'with the guide
vane member in different'settings and with the complementary
mechanical gear connected to obtain the desired increase in
speed of the turbine member. In accordance with this method,
braking torque is regulated by setting the engine speed and
thus the speed of the primary side of the torque converter.
Setting of the engine speed either requires utllizing the
engine to provide compression braking or utilizing an extra
gear pump driven by the primary side of the torque converter
to modulate the torque absorption thereof.
The setting of the different brakes and couplings,
and the regulation of the engine speed to provide a desired
braking capacity is preferably controlled by an automatic
control system controlling the transmissions. This system is
disclosed ln commonly assigned, copending Canadian application
Serial No. 318,972 entitled "Vehicle Transmission Control
System", filed on January 2, 1979.
As stated above, the guide vane member of the torque
converter is kept stationary to obtain a higher braking torque.
On the other hand, in another mode, the guide vane member is
released to rotate freely in either direction. Preferably, the
engine is made to absorb driving torque applied to the pump
member of the torque converter by the turbine member so as to
obtain a desired ratio between the engine speed and turbine
member speed. ~oreover, the engine can be set at a higher
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speed than that obtained when driven by the pump member of the
torque converter so as to reduce the braking torque.
In one embodiment, the primary side of the torque
converter drives a pump unitfor exertingbraking torqueon this
primary side of the converter and the torque is regulated by
varying throttling of the pump unit. Ina relatedembodiment, the pump
unit is of the variable capacity type and the torque is
regulated by adjusting capacity of the pump. In addition, the
torque can also be regulated by varying the throttling of and
varying the capacii~i of the pump unit.
Advantageously, the working fluid of the torque
converter is utilized in the pump unit driven from the primary side
of the torqueconverter andthe fluid passingthrough the pump unit is
channeled to circulate through a heat exchanger of the torque
,onverter. In addition, a part of the pressure fluid of the pump
~nit can be utilized to drive a cooling fan of the engine by
a hydrostatic motor.
Regarding the electronic control system referred to
above, the braking capacity level in relation to speed of the
vehicIe is preferably related to a manual setting and the
operation of the electronic control system for the transmission.
The control system is also used to control both connection of the
gear ratioin thecomplementary mechanicalgear andthe brakingor driving
of the engine to obtain the desired level of braking at different
vehicle speeds. Moreovex, where the torque converter has a
releasable pump member or turbine member, the electronic control
system isused-tocontrol releaseof thepump memberor turbinemember from
driving connection with the engine during the change of mechanical
gear ratios to obtain a hydraulic synchronizing effect to assist
in determination of
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the new engine speed.
In one mode of operation, the engine speed is adjusted
in relation to the mechanical gear shift by controlling the
fuel injection to the engine. Also, reverse gear can be connected
at low speed and the engine speed regulated to obtain braking to
a complete stop.
Other features and advantages of the invention will be
set forth in, or apparent from, the detailed description of the
preferred embodimenl:s found hereinbelow.
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Brief Description of the Drawings
Figure 1 is a longitudinal cross section of a trans-
mission in which, in a preferred embodiment, the braking method
of the invention is incorporated;
Figure 2 is a schematic diagram of the essential com-
ponents of the transmission of Figure 1 together with the mechanical
and electronic controls thereof; and
Figures 3, 4 and 5 are diagrams of selected operating
characteristics of a transmission incorporating the electronic
control system of the invention, wherein Figure 3 is a maximum
power tractive effort diagram showing operating conditions of the
transmission when the engine is applying tractive effort, and
Figures 4 and 5 show operating conditions of the transmission
utilizing the braking method of the invention, Figure 4 showing
the braking force that can be obtained at different speeds and
Figure 5 showing the relationship between braking force and engine
speed and the necessary braking force on the engine to provide a
dèsired braking force.
Detailed Description of the Preferred Embodiments
Referring now to the drawings, like elements are repre-
sented by like numerals throughout the several views.
Figure 1 illustrates a hydromechanical transmission
with which the vehicle transmission braking system of the invention
can be used. ~t the left end of Figure 1 there is shown a torque
converter TC including a rotating casing 2 adapted to be driven by
a vehicle engine or the like via abutment means 2a. Internally,
the illustrated torque converter comprises a pump member 3 having
a ring of pump blades 4 mounted thereon. The torque converter
further comprises a turbine member 5 having a ring of turbine
blades 6 mounted thereon and a guide vane 7 having a ring of guide
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blades 8 thereon, wherein said guide blades may be used as a turbine.
Connected to the turbine member 5 is a hub 14 to the outer periphery
of whi~h is attached a friction disc 12. The rotating casing 2
includes an inward extension 2b located between the disc 12 and the
pump member 3 and a servo piston 10 on the outer side of disc 12.
The torque converter shown herein is of the releasable pump member
type which is shown and described in detail in prior U.S. Patent
No. 3!893,551, issued July 8, 1975. In accordance therewith, the
pump member 3 is movable to the left to engage the pump member 3
with the rotating casing at conical friction coupling 9 for hydraulic
drive. In another mode of operation, pump member 3 is moved to the
right, releasing coupling 9, and the servo piston 10 is actuated to
urge disc 12 into frictional engagement with extension 2b for
direct drive between the rotating casing 2 and the turbine member 5.
The turbine member 5 and the hub 14 are drivingly engaged with the
turbine shaft 16. The guide member 7 is mounted on a guide member
shaft 18 which rotatesrelative to turbine shaft 16 and which is
mounted on the stationary portion of the casing at bearings 20.
Shaft 18 is connected via a hub and friction discs to a brake 26
operable by servo piston 23 for holding the guide member 7 stationary
for "single rotation". Shaft 18 is further connected to a planetary
gear 22, the carrier of which is connected to friction discs form-
ing a part of brake 24 which is operated by servo piston 25, whereby
the guide member rotates oppositely from the turbine member for
"double rotation". Arrangements for braking the guide member for
single or double rotation are well known, one example being shown
in the prior U.S. Patent No. 4,010,660, issued March 8, 1977. In
hydraulic drive, torque multiplication is provided via the guide
member blades, and the output of increased torque via the turbine
member to the turbine shaft 16. Double rotation with brake 24
actuated allows a much higher multiplication of torque, but over
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a smaller range of speed ratios, than does single rotation (engage-
ment of brake 23) wherein speed ratio is defined as the ratio
of turbine shaft speed to rotating casing speed. Torque
~ultiplication decreases with increasing speed until it becomes
advantageous to disconnect hydraulic drive, i.e. disconnect the
conical coupling 9, and to actuate servo piston 10 to drive the
turbine shaft 16 directly from the rotating casing 2 via elements
12 and 14.
The torque converted includes a heat exchanger 68
through which fluid is pumped by means of a pump unit 71
operated by means of a gear 70 via an intermediate gear 72. A
system including, in a torque converter, a heat exchanger of this
type, together with a pump unit and the appropriate fluid lines,
is shown in greater detail in prior U.S. Patents 4,056,019 and
4,058,980, issued respectively on November 1, 1977 and November
22, 1977. O
To obtain overspeed of the turbine, there is provided
downstream from the torque converter, a mechanical gear trans-
mission comprising a first portion P having four forward gear
ratios and a reverse gear, and a second portion, R referred to as
a "range gear" having either a 1:1 drive or a further gear
reduction. A transmission having such a first portion P followed
by a second "range" portion R is shown in greater detail in commonly
owned copending Canadian Application Serial No. 289,475, filed
October 25, 1977.
The turbine shaft 16 is connected to a ring gear 30.
The secondary or output shaft of this first portion is designated
as 32. Ring gear 30 drives a planetary gear 33 having a plura-
lity of sections including a large diameter section 31 splined
onto an intermediate diameter section 34 with a smaller diameter
section 35 to the right. Intermediate between sections 34 and 35
is a bearing means for mounting this planetary gear 33. Sections
31, 34 and 35 are respectively drivingly engaged with splines of sun
1~18318
gears 36, 38 and 40. Section 35 is further engaged with a ring
gear 50 operating as a reverse gear. Sun gears 36, 38 and 40 are
either released for free rotation or connected to the stationary
portion of the casing via friction brakes 46, 44 and 42, respec-
tively, which friction brakes are actuated by servo pistons 47,
45 and 43, respectively. Reverse gear 50 is selectively engaged
with the casing via friction brake 50 which is actuated via servo
piston 53.
Alternatively, ring gear 30 can be connected directly
to the carrier of planetary gea~ 33 and hence directly to secondary
shaft 32 by engagement of friction clutch 48, the latter caused by
actuation of servo piston 49, this in turn urging member 51a to
the left to turn lever 51b such that its upper portion moves to the
right to engage clutch 48.
Shaft 32 extends toward the right in Figure 1 into the
second portion or "range gear" whereat it is drivingly engaged
with an elongated splined member 54 which is drivingly engaged with
both a hub 56 and planetary gears 64. The holder 60 of planetary
gear 64 is drivingly engaged with a secondary gear 58 which is
the output shaft of the entire transmission. Planetary gear 64 is
engaged with a ring gear 66 which can be braked relative to the
stationary portion of the housing by means of a friction brake 67
which is actuated by servo piston 69. This would permit a speed
reduction between shafts 32 and 58. Alternatively, shafts 32 and
58 may be operatively engaged to each other via hub 56 and friction
clutch 62, the latter frictionally engaginghub 56 with the planetary
gear holder 60. Friction clutch 62 is actuated via a servo piston
63 which acts via a lever system 65.
Referring now to Figure 2, there is shown an electro-
a hydraulic-electronic control system described in the aforementioned
' copending'application Serial No. 3~ L , entitled "Vehicle
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Transmission Control System". In Figure 2, mechanical conncctions
are indicated in solid lines and electrical connections in dashed
lines.
For convenience, the hydraulic systems contained within
Figure 2 will be described first. A pump system 300 includes a
high pressure gear pump GPH and a pair of low pressure gear pumps
GPL, the pressure of which is controlled by a solenoid valve CBV.
There is also included aheat exchanger HE as described in the
previously mentioned U.S. Patent No. 4,058,980. These pumps pro-
vide the pressurized oil to operate the valves of the system and
the pressurized oil which flows through the valves to the various
servo pistons and to the torque converter chamber. The system com-
prises a first valve 302 which controls the flow of fluid to the
torque converter, a second valve 304 which controls the flow of
flui~ to single and double rotation servo piston 23 and 25, a third
valve 306, which together with secondary valves 306A and 306B,
controls the flow of fluid to the first portion of the mechanical
transmission and lastly a valve 308 which controls the flow of
; fluid to the servo pistons 63 P~d 69 of the range gear. Oil under
pressure is delivered from pump system 300 to the valve 302 for
delivery to the torque converter for selecting direct or
hydraulic drive, and to the valve 304 for delivery to single and
doubIe rotation servo pistons 23 and 25. Oil under pressure is
also delivered from pump system 300 to the valve 306 and its
secondary valves 306A and 306s for delivery to the servo pistons
of the first portion P of the mechanical transmission and to
valve 308 for delivery to the servo pistons of the second portion
R of the mechanical transmission. Finally, low pressure fluid
from pump system 300 is also delivered to all the electro-
hydraulic solenoid valves for controlling operation of said
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valves 302, 304, 306, 306A, 306B and 308.
Referring again to Figure 2, and specifically to
valve 302, it will be seen that the pressurized oil enters the
valve at line 70'. With the spool of valve 302 in its neutral
position, the
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1118318
torque converter is in its neutral position with neither the couplin~
9 nor the disc 12 engaged with extension 2b of the rotating casing 2.
Movement of valve 302 in one direction will then connect the pres-
surized fluid from line 70~to line H for hydraulic drive and move-
ment of this valve in the other direction will connect such pres-
surized fluid with line D for actuation of servo piston 10 and
hence dlrect drive. It is obvious, therefore, that one cannot
place both lines H and D under pressure at the same time.
Turning to valve 304, pressurized fluid through line
71 will flow through either a first line SR or a second line DR,
depending on the direction of movement of valve 304, to actuate
either single rotation piston 23 or double rotation piston 25.
At valve 306, pressurized oil entering at line 74 is
delivered either through line R to servo piston 53 or through line
S to the two further valves 306A and 306B. Valve 306A has three
positions including two end positions whereat the entering pres-
surized fluid is delivered to either servo piston 45 or servo piston
43. The third position is a neutral position whereat the fluid
passes through valve 306A to valve 306B. The latter, in ~Urn,
has two positions, a first and position whereat this pressurized
fluid passes through a line to servo piston 47 and a second posi-
tion whereat this fluid passes through another line to servo piston
49.
Finally, valve 308 receives pressurized fluid from line
76. As this valve 308 is moved to its end positions this pressur-
ized fluid is delivered to either servo piston 63 which operates
friction clutch 62 or servo piston 69 which operates brake ~.
Thus, in summary, the hydraulic control valve system,
including valves 302, 304, 306, 306A, 306B and 308, controls the
flow of oil to the servo-pistons which directly engage the various
brakes and clutches in the transmission of Figure 1 as described
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above, with pressurized oil being obtained from thefeeder pumpsystem
300, which is driven by the primary side of the transmission.
The hyclraulic valve system is, in turn, controlled by means of
solenoid valves which are described below and which, through
electrical signals, control the flow of oil actuating the various
servo-pistons in the hydraulic valve system.
The system of Figure 2 further includes an engine brake
cylinder CEB, a fuel injection cylinder CFI and a fuel cut-off cyl-
inder CFC which are controlled by solenoids EBV, FIV, FCV. These
operators and their functions are conventional.
As mentioned hereinabove and shown in Figure 2, the
setting of the transmission is determined by plurality of solenoid
-type valves. These valves are indicated in Figure 2 at DV, HV,
SRV, DRV, FV, RV, EHVl to EHV6 and CBV. These valves control, via
the hydraulic valve system including valves 302, 304, 306, 306A,
306B and 308, the flow of oil for connection of direct drive, hy-
draulic drive, single rotation drive, double rotation drive, forward,
reverse, mechanicalgears oneto eight,and thepressure andcapacitor pump
system 3respectively. In particular,solenoid valvesDV andHV controlval
ve 302,solenoid valvesSRV and DRV control valve 304, solenoid valves
F and R control valve 306, solenoid valves EHVl to EHV6 control
valves 306A, 306B and 308, while, as mentioned previously, solenoid
valve ~BV controls pump system~300. In addition,the furthers~lenoid
valves FIV, FCV, and EBV control the engine-influencing devices for
fuel injection, fuel cut-off, and engine braking, respectively,
as mentioned hereinabove and as discussed in more detail below in
connection with the present invention.
Turning again to Figure 2, the input signals referred
to above comprise shaft speed signals which appear on lines 309,
310 and 311, throttle position signals which appear on lines 312,
313 and 314, brake pedal and handbrake signals which appear on lines
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315 and 316 respectively, oil level and temperature safety signals
which appear on lines 317 and 318, respectively; selector lever
~osition signals which appear on lines 319, 320, 321, 322, 323 and
324; and brake lever position signals which appear on lines 325,
326 and 327.
The shaft speed signals are square wave, TTL pulses con-
sisting of two levels, viz., 0V and +5V. The pulses are obtained
from the sensor/amplifier units Gl, ~2 and G3 disposed adjacent to
gear teeth rotating with the engine shaft ES, converter turbine
shaft CTS, and transmission output shaft TOS, respectively, as il-
lustrated.
Considering the other input signals in more detail, the
throttle position signals appearing on lines 312 to 314 are related
to the position of the throttle lever indicated at 330 and these
signals include a variable voltage between 0 and 5V which is, pro-
portional to the throttle position and which is provided bv a po-
tentiometer (not shown), the tap of which is attached to the throttlc
lever 330. Two further signals, indicating the terminal positions
"throttle released" (or N) and "kickdown" (or KD), respectively,
are obtained by use of contactors 332 and 334 which open or close
connections to ground. When the contactors 332, 334 are in the
open positions, the microcomputer holds the signal line at +5V~
The brake pedal (Br) and handbrake (HB) signals on
lines 315 and 316 are obtained in a same manner, i.e., through the
use of contactors, indicated generally at 335 and 338, respectively,
which provide openings or closing of a connection to ground.
The oil level (OL) and temperature safety (TS) switch
signals appearing on lines 317 and 318 are provided by switches,
indicated generally at 340 and 342, and are held at ~24V (the battery
voltage) during normal operation. Excessively low oil level or high
oil temperature cause the respective switch 340 or 342 to close a
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33:~8
connection to ground, thus lighting a warning lamp WL on the instru-
ment panel and simultaneously activating a delay circuit in the
microcomputer. This delay circuit allows time for the driver to take
some independent action before the microcomputer releases the trans-
mission so as to prevent damage.
The selector lever and the brake lever are intended to
be directly controlled by the driver through the microcomputer200. The
signals are produced, as shown, by a number of switch contactors,
which control the completion of connections to ground, thereby pro-
viding for a combination of signals. The microcomputer 200 holds
the signal lines at +5V for an open connection. The selector lever
is indicated in Figure 2 in dashed lines at 344 and the contact
plate at 346, while the six output lines 319 to 323 are respectively
dedicated to the following driving settings: reverse (R), reverse
neutral ~RN), neutral (N), forward neutral (FN), forward (F), low (L)
and extra low (EL). In the neutral (N) position, the transmission
brakes arereleased and the turbine pump is released. In the forward
neutral (FN) position, the DR brake (or possibly the SR brake) is
applied. The turbine pump is released, so that free wheeling is
provided and instantaneous vehicle stopping can be provided. In
the forward (F) position,theturbine pumpmember isengaged andthis is
the normal driving position. The other positions are self-explana-
tory. It will, of course, be understood that more and different
settings can be provided as desired.
- The brake lever, which is indicated at 348, uses lines
325 to 327 to indicate eight different braking levels by virtue
of the pattern of switch contactors 348a illustrated. These brake
lever input lines like those from the selector lever are connected
to the microcomputer 200.
Again, it will be understood that the foregoing listing
is not exhaustive, and the microcomputer 200 has a capacity to
process many more of each of the different types of signals discussec?
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1118318
The systcm is powered by the vehiclc battery B which
provides ~24V and the system ground. The battery B is connected to
a voltage regulator 352 which provides a stabilized +5V supply,
the voltage regulator being located in the selector lever box or
housing with an ignition switch 350 and serving to supply the micro-
computer, the shaft speed signal amplifiers, and the throttle posi~
tion potentimeter mentioned above. All connections to ground in
the system axe made through a common ground line, connected to the
minus pole of battery B via the microcomputer 200.
Before discussing the present invention in more detail,
the operating characteristics of the overall system will be briefly
considered. The full throttle performance characteristics of an
engine-transmission unit as described above are shown in Figure 3,
including the characteristics for double rotation drive, single
rotation drive, and direct drive together with the eight-speed gear.
For sake of simplicity, the hydraulic drive characteristics are
shown for the first four mechanical gear ratios only; however, all
combinations are, of course, possible and the determination and
control of these combinations (for all throttle positions) neces-
sitates the use of an electronic control system such as described in
the above-mentioned copending application. In addition, the deter-
mination and control of transmission settings (including the control
of the engine) during braking by overspeeding the turbine or varia-
tions thereof,as provided in accordance with the invention,generally
require such a control system. In brief, braking by means of over-
speeding the turbine is obtained for the transmission of Figure ]
by releasing the direct clutch friction disc 12,and connectingcouplin
9 ofthe releasablepunlp member3 andconnecting thesingle rotationbrake
2~ when the speed ratio as defined above is greater than unity.
This condition is achieved by providing for a gear ratio in the
mechanical gear which is lower than wnat wo~ld normally be
provided. With this
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lllB318
technique, the direction of the flow of power in the transmission
is thus reversed, thereby producing a braking or retarding effect.
The characteristics of the retarding function are shown in Figures
4 and 5, where regulation or modulation of the amount of braking
is achieved in part by influencing or controlling the engine, by
means of pressurized air devices, vacuum devices, electro-magnet
devices, or other electrical devices, which devices are energized
by means of electrical signals. Figure 5 shows other variations of
hydraulic braking, where either the pump member 3 or the single
rotation brake 26 is released, that is, free to rotate. These
variations also produce a retarding effect which is modulated by
influencing or controlling the engine as described in more detail
hereinafter.
It will be understood that mechanical transmission P
and R provide eight forward gear steps as shown schematically and
numerically in Figure 3. The first four gear steps are of course
the four forward gear ratios in mechanical transmission P with the
brake ~ of the mechanical range gear R connected to the stationary
O ...
casing, thereby providing a reduction gear ratio through the mech-
anical transmission R. The next four gear steps V through VIII
again comprise the same four gear steps of mechanical transmission
P, but this time with the input to mechanical range gear trans-
mission R connected directly to the output shaft 58 via engagement
of brake 62.
Depending on how hard the throttle is pressed, one will
have different tractive efforts. If the throttle is pressed down
to maximum, thenOne achieves the tractive effort illustrated in
Figure 3 at curvePn/I However, normally, the vehicle accelerates
faster than the engine in the low gears, and therefore this high
tractive effort is not actually obtained except when climbing
extremely high grades. When, however, the vehicle has accelerated
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to a certain point in relation to the engine speed, then the
guide member 7 is disconnected from the turbine and connected to
the stationary casing, i.e. brake 24 is released and brake 26 is
engaged, which of course comprises normal single rotation drive.
This condition remains until the point is reached whereat direct
drive is required at which point coupling 9 is disengaged, freeing
the pump member 3 from the rotating casing an ~piston 10 is then
activated to enqaqe the disc 12 aqainst the extension 2b of the
rotatinq casinq 2. The point at which the transition from one
condition to the other occurs is related to the throttle-pedal
position which will be at different speed ratios between the pump
member and the turbine member after the vehicle has accelerated
sufficiently. In first gear the vehicle can now accelerate up to
about 12 km/hour and the tractive effort is represented by the curve
Pd/I in the case of maximum throttle. Normally the first two
or three gear steps of the mechanical transmission P and R are
used only for starting un~er severe conditions or for driving
fully loaded up very high grades. Normally, therefore, the automatic
control means may have already connected up to the fifth gear or
?ossibly up to the eighthgear before there arises the need for
applying some type of braking or retardation.
In conventional hydraulic braking, at the torque conver-
ter, the direct drive is connected and the guide member 7 is held
fixed to the stationary casing or at lower vehicle speeds the
guide member 7 can be connected to the turbine at brake 24 (again,
with direct drive connected) thereby making the guide member rotate
backwards.
While this conventional type of braking is satisfactory,
it does not provide the ability to regulate or modulate the braking
except by connecting different gears. Therefore, in lieu of this
conventional braking, in accordance with the present invention it is
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possible to provide a hydraulic braking by overspeeding the turbine,
having disconnected the direct drive connection. According to this
arrangement, the automatic control means must connect-a gear for
a certain overspeeding of the turbine as shown diagrammatically
in Figures 4 and 5. In Figure 4 the lines marked n with the indices
I-VIII indicate turbine speeds, and the areas marked P with indices
I-VII indicate the retardation or braking force obtainable, the
lower limits of the obtained retardation force being with a released
guide membex and with the engine running due to torque transmitted
from the turbine member to the pump member. These lower limits
of Figure 4 are represented in Figure 5 by the lines150 (Ph I-IIIc')
for three different gear ratios, and the upper lines in Figure 4
are represented in Figure 5 by the lines marked 158 (Ph I-IIIa').
There is a difference, however, in Figure 5 wherein the lines re-
late to constant braking torque on the engine by compression braking
or the like and the speeds of the engine in Figure 5 are in accor-
dance with the dot-dash lines marked 152, 156 and 160 (nlI-IIIa-c).
Figure 5 is the more theoretical diagram while Figure 4 shows the
limits of retardation force obtainable by controlling the connection
of gear ratios in relation to speed and engine speed also taking
into consideration the temperature of the transmission, etc.
As will be seen from Figure 4, seven braking fields
(i.e., defined areas in drawings) are obtained with eight gear
steps. If only four gear steps are provided, only three braking
fields are obtained. Figure 5 illustrates the braking capacity
provided with a four-speed complementary mechanical gear affording
three braking areas. In Figure 5, the curve 150 illustrates the
braking capacity with a released guide vane member and the turbine
member connected to a third gear, of the mechanical gear,with a
maximum vehicle speed of 66.5 km/hour. This braking corresponds to
the engine speed in accordance with curve 152 which is the speed
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1~18318
the engine obtains by the transferred torque through the torque
converter. To obtain braking according to curve 154 it is nec-
essary to reduce theengine speed. The line 154 also illustrates
the braking obtained with the guide vane member stationary wherein
the speed of the2 WDS engine produced by the transferred torque is
represented by curve 156. Curve 158 illustrates the braking capa-
city when the engine speed is reduced to the speed shown by curve
160. A vehicle speed of 66.5 km/hour in the third gear corresponds
to a turbine member speed of 3500 rpm. At speeds between 48 and 52
km/hour it is appropriate to change the mechanical gear to second
gear whereby the same circumstances obtain as described with
respect to third gear, and when shifting the turbine member to be
connected to first gear between 34 to 38 km/hour the field of
braking marked III in Figure 4 is obtained as shown.
As described abovein connection with Figure 5, the
hydraulic braking capacity is to be modulated by varying engine
speed. Further, under certain circumstances the braking is decreas-
ed by driving the engine utilizing the fuel control unit formed
by the fuel injection cylinder CFI and fuel cut-off cylinder CFC
shown in Figure 2. In addition to the increased braking effect,
under certain circumstances a certain reduction of the engine speed
must be effected, the engine speed being dependent on the torque
transferred from the rotating turbine member to the pump member of
the torque converter. An arrangement for providing this reduction
in engine speed is shown in Figure 2 in the form of engine brake
cylinder CEB. The function of the brake cylinder CEB is to control
throttling of the exhaust pipe of the engine. The braking effort
on the engine is small in relation to the braking effort of the
turbine member. Actually, the turbine member, when overspeeded,
operates as a pump delivering energy to the pump member of the
hydraulic torque converter which is then operating as a turbine.
~r~, - 20 -
1~183~8
Naturally the efficiency when thus inverting the function of the
torque converter is very low and therfore the drivin~ torque on
the engine is relatively small. However, as stated, this is suf-
ficiently high to drive up the engine to a high speed relative to
the turbine speed at least when the turbine speed is higher than
normal maximum speed.
It will be understood that cooling does not present any
problems and, in fact, this cooling will be taken care of even when
the engine is kept at low speed. However, cooling of the torque
converter can present problems. Because of this the torque con-
verter is supplied with an extra pump located, for
instance, on one of the power take-offs. This pump will be dis-
connected during normal driving. This pump will then receive fluid
from the secondary side of the heat exchanger HE and deliver fluid
back to the primary side of the heat exchanger HE and/or be used
/of
to increase the circulation fluid through the torque converter to
obtain cooling of the torque converter and to increase the cooling
of the oil. Under these circumstances it may be desirable to use
a part of the oil to drive the engine fan at a higher speed.
Another solution of the problem discussed above is to
only utilize the pump unitdriven bythe power take-off for braking
the engine, increasing the fluid through torque converter and drivin~
the fan of the engine. For this purpose it is appropriate to util-
ize an adjustable vane pump. However, even if this pump can be set
at zero capacity, it must be possible to disconnect it with a re-
leased clutch during normal operation.
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