Hybrid gearbox

A hybrid gearbox for land vehicles and automotive machines is designed such that the starting up and maneuvering operations of the vehicles are powered by a hydromotor and their normal travel is powered by a direct coupling with an internal combustion engine. In its normal operation, the internal combustion engine drives a neutrally positioned axial piston pump as well as the driven shaft. When operated under load, the driven shaft is decoupled from the internal combustion engine, and the latter drives the variable axial piston pump which in turn delivers pressure to the hydromotor.

FIELD OF THE INVENTION 
The present invention relates to a hybrid gearbox for all kinds of land 
vehicles. The new gearbox is particularly suited for vehicles which 
frequently change their travel direction, which frequently start up, which 
exhibit large tractive powers and/or low travel speeds and high final 
speeds at the same time. 
BACKGROUND OF THE INVENTION 
Vehicles of this type are in particular construction machines such as wheel 
loaders, V-dump cars (dumpers), motor graders or back hoe loaders 
(multi-purpose dredges). 
In the field of forestry equipment, the hybrid gearbox is suitable for 
skidders, forwarders, harvesters and lumber transport vehicles for use in 
mountainous terrain. 
In the field of municipal vehicles, the hybrid gearbox can be used for 
street sweepers, snow removers, rotary snow plows, lateral mowers for lane 
edges and side street maintenance vehicles. 
The hybrid gearbox is moreover suitable for use in city buses which have 
their engine built in transversally to the vehicle, in a horizontal 
version and in connection with an axle without bevel differential gear. 
SUMMARY OF THE INVENTION 
It is the object of the present invention to provide an as simple as 
possible gearbox for a hybrid drive by means of which the land vehicle is 
driven either directly or hydraulically by an internal combustion engine, 
in particular a diesel engine. 
This object is accomplished in that the hybrid gearbox is either directly 
mounted on the flywheel casing of the internal combustion engine or that 
it is driven via a shaft. By means of an automatic transmission, the 
internal combustion engine may be directly coupled to a driven shaft or a 
hydromotor which in turn receives pressure from an axial piston pump, said 
pump being likewise driven by the internal combustion engine. Further 
features may be gathered from the following description and the attached 
claims. 
From the input shaft of the gearbox, the hydrostatic pump with variable 
swept volume is driven via an intermediate gear. By selecting this 
transmission ratio appropriately, a speed of the combustion engine which 
has been kept low so as to reduce noise and emissions can be increased to 
a higher value, thus allowing the use of smaller, lighter and cheaper 
pumps. Via a further intermediate gear, e.g. for bridging the axle 
distance required for all-wheel drive, a gearwheel is driven which is 
supported on the driven shaft of the gearbox. This gearwheel may be 
connected to the drive shaft by means of an axially movable coupling 
sleeve in such a way that it remains in rotation but also remains 
slidable. 
Furthermore, a hydromotor is mounted on the gearbox and is in turn 
connected to a second, similarly supported gearwheel on the driven shaft. 
An intermediate gear may be provided in the form of a block gear for 
generating a larger transmission ratio between the hydromotor shaft and 
the driven shaft. 
The direct connection of the drive shaft allows auxiliary drives to get a 
movement at the full engine power. Depending on the space available, the 
hydraulic pump and the hydromotor may either both be mounted on the drive 
or driven side of the gearbox, or, alternatively, one on each side. 
Further details, features and advantages of the invention may be gathered 
from the following description of an embodiment which is schematically 
shown in the drawing.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENT 
An input shaft 20 is driven by an internal combustion engine (not shown) at 
22, either directly or via an intermediate gear. A toothed gear 24 is 
connected to said drive shaft 20 in such a way that it remains in rotation 
said toothed gear 24 meshing on the one hand with a pinion 26 on a shaft 
28 and on the other hand meshes with an intermediate gear 30 on a 
countershaft 32. In the present example, provided with the input shaft 20 
is a connection 33 for a suggested power take-off shaft 35. In the case of 
multi-purpose vehicles or machines, this would serve for flanging an 
aggregate (35) thereto. 
Shaft 28 drives a first pump 34 for generating a hydrostatic pressure as 
well as a second pump 36 for moving a control piston 40 in a cylinder 42. 
In the view of FIG. 1, the intermediate gear 30 engages a gearwheel 50 
which is mounted on a driven shaft 52 which latter is connected to the 
vehicle drive wheels. Consequently, shaft 52 as shown in the Figure may 
have a flange 54 on its left which transmits the driving moment to the 
front axle, as indicated by arrow 56, and may exhibit a flange 58 on its 
right which is connected to the rear axle, as indicated by arrow 60. 
In the position marked in the drawing the control piston 40 is in its left 
end position in the cylinder 42. This also causes a piston rod 62 to be 
moved to its left position with a coupling finger 64 and a coupling claw 
66. In this position, the coupling claw 66 has been moved to the position 
indicated by continuous lines, in which the gearwheel 50 meshes with the 
intermediate gear 30, thus causing the driving torque to be transmitted to 
the driving wheels at a ratio of almost 1:1 or geared down to a small 
ratio. In this position of the coupling claw 66, which is coupled to shaft 
52, engages the gearwheel 50, and a further gearwheel 70 does not engage 
coupling claw 66, so gearwheel 70 can rotate on bearing 134 around shaft 
52 in response to gear rim 72 on a shaft mounted below which may be driven 
by a hydromotor 80. This driving power from the hydromotor 80 will not be 
transmitted to the driven shaft 52 as long as the gearwheel 70 and 
coupling claw 66 do not mesh. 
Only after the control piston 40 in cylinder 42 has been moved to the 
right-hand position and the coupling finger 62 has moved the coupling claw 
66 into the position indicated by broken lines will the gearwheel 50 
disengage the coupling claw 66 and rotate on bearing 132 in response to 
the intermediate gear 30, and the gearwheel 70 will mesh with the coupling 
claw 66. The driven shaft 52 will thus be driven by the hydromotor 80. 
This known adjustability of axial piston machines, however, may also be 
made a feature of the hydromotor 80. For this purpose, a control line 84 
is provided. The hydrostatic motor 80 is used for starting up. By 
selecting the appropriate setting for the piston in the cylinder 42, 
namely by admitting pressure from the second pump 36 to act upon the 
piston via a line 88, the coupling claw 66 is in its above-described 
right-hand position and the driven shaft 52 is coupled to the pinion 72 
through gearwheel 70, while the gearwheel 50 runs free in this position. 
Only after switch-over to mechanical operation, as set out hereinafter, 
will the piston 40 be acted upon by pressure from a pressure line 86. 
This known adjustability of axial piston machines, however, may also be 
made a feature of the hydromotor 80. For this purpose, a control line 84 
is provided. The hydrostatic motor 80 is used for starting up. By 
selecting the appropriate setting for the piston in the cylinder 42, 
namely by admitting pressure from the second pump 3 to act upon the piston 
via a line 88, the coupling claw 66 is in its above-described right-hand 
position and the gearwheel 70 on the driven shaft 52 meshes with the 
pinion 72, while the gearwheel 50 runs free in this position. Only after 
switch-over to mechanical operation, as set out hereinafter, will the 
piston 40 be acted upon by pressure from a pressure line 86. 
Swinging out the variable hydraulic pump 34 allows the drive to be engaged 
without any tear and wear, and the vehicle will accelerate up to a speed 
which is predetermined by the design of the hydrostatic drive. The 
direction in which the pump 34 is swung out also allows for determining 
the travel direction. 
For switching over from a hydrostatic drive to a purely mechanical one, the 
hydrostatic drive is adjusted to be torque-free by suitable pump control 
via a transmission management box (TMB), and at the same time the diesel 
engine is accelerated to such a speed that the rotational speed of the 
mechanical gearwheel 50 is identical with the speed of the drive shaft as 
defined by the present travel speed. These speeds are measured by means of 
respective sensors 100, 102, 104, which transmit their respective signals 
to the TMB. The speed of the hydromotor 80 is determined by sensor 100, 
the speed of the driven shaft 52 is determined by sensor 102 and the 
reduced speed of the internal combustion engine is determined by sensor 
104, and all data is delivered to the TMB via the dot-dash lines. The 
positions of the piston 40 and of the piston rod 62 are traced by the 
sensors 94 and transmitted to the TMB via lines 96. During automatic 
operation, the TMB automatically controls the control of the injection 
pump of the internal combustion engine via a line 98. 
The shift point position, i.e. the position at which automatic switch-over 
from the starting up drive by means of the hydromotor 80 to the direct 
drive is desired, can be preselected via lever 110. The changeover from 
"automatic (112)" to "hydrostatic (114)" with the options of 
"forward"/"backward", or to "mechanical (116)" is effected at the switch 
box 120. The travel speed is at any rate controlled in the usual manner 
via pedal 122. 
It is useful to switch between hydraulic and mechanical operation at the 
engine speed at which the torque of the internal combustion engine is 
highest. However, this switching time may also be set to any torque above 
the idling speed if the torque of the internal combustion engine is 
sufficient to further accelerate the vehicle. 
This measure allows for all advantages of the hydrostatic travelling gear 
such as 
continuous control of the travel speed independent of the engine speed of 
the internal combustion engine, 
starting up and change of the travel direction without any tear and wear 
favourable efficiency as compared to hydrodynamic torque converters 
to be exploited in the lower travel speed range. 
The known worse efficiencies of the hydrostatic travelling gear at larger 
control ranges of the speed, due to small swinging angles of the 
hydromotors and large swept volumes of the pump and corresponding power 
losses, are avoided by the mechanical drive. Switch-over from hydrostatic 
to mechanical drive may either be directly effected by the driver or be 
controlled by an automatic transmission in the TMB, i.e. independently of 
the driver. 
If a power shift element is additionally provided between the hybrid 
gearbox and the hydromotor (cf. DE 44 14 127, C1), a conversion range is 
obtained e.g. for a wheel loader which allows for travel speeds from 0-25 
km/h to be obtained with small hydrostatic units with the high transverse 
forces for loosening the material, at the same time enabling a travel 
speed on public streets or roads of up to 50 or 62 km/h without causing 
essential costs. 
As a side effect, in vehicles for which a retarder is mandatory by law, the 
hydrostatic pump can be used as a tear- and wear-free retarder for the 
mechanically operated speeds directing it against a pressure control 
valve. The heat generated is discharged to the environment via the 
radiator which is provided anyway. 
This is to be demonstrated with the concrete example of a lumber transport 
vehicle. 
______________________________________ 
Dead weight 13 t 
Useful load 12 t 
Total weight 25 t 
Tires 700-22.5 
Tire radius 0.53 m 
Tractive power at least 15 t 
Rolling resistance (firm street or road) 
0.03 
Mechanical efficiency 85% 
Drag coefficient c.sub.w 1.0 
Cross-sectional area 9.1 m.sup.2 
Axle transmission ration 13 
Diesel engine 128 kw (175 
HP) at 2,500 
rpm 
Maximum torque 650 Nm at 
1,200 rpm 
Transmission ratio 0, 78 
for mechanical drive 
Transmission ratio for hydrostatic drive 
1st gear 7.96 
2nd gear 2, 47 
Hydraulic pump, q max. 71 cm.sup.3 /rev. 
Transmission ratio between diesel and pump 
0.75 
Maximum pump speed 3,300 rpm 
Maximum pump swept volume 
235 1/min 
Hydromotor, q max. 160 cm.sup.3 per 
revolution 
Hydromotor, q min. 55 cm.sup.3 per 
revolution 
______________________________________ 
The following values are obtained based on the above main data: 
______________________________________ 
Maximum hydromotor speed 
4,000 rpm 
Maximum hydromotor torque 
1,020 Nm at q max. and 
420 bar 
Maximum travel speed 
in the 1st hydrostatic gear 
7.85 km/h 
in the 2nd hydrostatic gear 
25.3 km/h 
Maximum tractive force 
in the 1st hydrostatic gear 
152,350 N 
in the 2nd hydrostatic gear 
47,275 N 
Maximum travel speed 
in the mechanical gear 
50.2 km/h 
Maximum tractive force 
at maximum travel speed 
8,000 Nm 
Maximum gradient which can be 
approx. 4.7% at 25 
travelled at maximum torque of the 
km/h 
diesel engine 
______________________________________ 
For switching with optimum exploitation of the acceleration, the 
hydrostatic drive is driven to its maximum speed in the 2nd hydrostatic 
gear, i.e. a hydromotor speed of 4,000 rpm which corresponds to a speed of 
about 1,620 rpm of the driven gear shaft. The speed of the diesel engine 
is stepped down or up until it likewise generates this speed of the idle 
wheel on the drive shaft, together with the mechanical transmission. Then 
a switch-over from the hydrostatic to the mechanical drive is effected. 
It will be obvious to those having skill in the art that many changes may 
be made to the details of the above-described embodiment of this invention 
without departing from the underlying principles thereof. The scope of the 
present invention should, therefore, be determined only by the following 
claims.