The invention relates to steering and traction control for electrically propelled vehicles, and particularly to differential steering and traction control systems for electrically propelled mowers.
Conventional riding mowing machines such as greens mowers, fairway mowers, and other wide-area mowers are used to mow grass on golf courses, athletic fields, and other areas that require high quality mowing. These mowers are steered solely by changing the angle of the rear wheel or wheels. The operator's steering wheel is mechanically or hydraulically connected to the rear wheel steering geometry.
The two front wheels on conventional riding mowing machines provide all, or the majority of, the vehicle's wheel torque. During turns, the torque of each powered wheel, either the left or right wheel, is passively adjusted so that the outside wheel turns faster than the inside wheel, in a ratio approximately proportional to the angle of the steered rear wheel. This method is relatively simple, but does not optimize traction. Additionally, if one of the powered wheels loses traction, all power from the propulsion drive of the riding mowing machine goes to the slipping wheel, and the machine may be stopped until traction returns.
A riding mowing machine is needed for golf courses, athletic fields, and similar applications that provides optimal traction. A riding mowing machine is needed that will continue to move even if one of the driven wheels loses traction.
Riding mowing machines with electric propulsion have been proposed. For example, U.S. Pat. No. 5,406,778 relates to an electric drive riding greens mower with a battery power source, an electric motor to provide driving torque, and electric motors for each of the cutting reel units.
The invention provides a differential steering and traction control system for a riding mowing machine.
While this invention is susceptible of embodiment in many different forms, there are shown in the drawings, and will be described herein in detail, specific embodiments thereof with the understanding that the present disclosure is to be considered as an exemplification of the principles of the invention and is not intended to limit the invention to the specific embodiments illustrated.
In one embodiment, riding mowing machine 100 may have a frame supported by a pair of front wheels 101, 102, each wheel powered by an electric motor wheel drive, and a single rear wheel 103 which may have an electric steering motor and optionally may include an electric motor wheel drive. The riding mowing machine may carry or support a plurality of cutting units 105 in a first row, and one or more cutting units 106 in a second row. Each cutting unit may be a reel-type cutting unit with a generally horizontally aligned reel, each reel powered by an electric motor. Alternatively, each cutting unit may be a blade rotating on a vertical shaft and covered by a deck.
In one embodiment shown in
In one embodiment shown in
Thus, electronic control unit 120 not only provides commands to an electric steer motor to turn rear wheel 103 in proportion to the steering wheel input, but also provides speed commands to the electric motor wheel drives to left and right wheels 101, 102 to continuously adjust their speed to remain in the correct proportion depending on the commanded speed and rear wheel angle.
In one embodiment shown in
In the computer algorithm shown in
In one embodiment, the mowing machine may not execute spin turns because the chassis of the mowing machine has 35.5 mm of caster in the steerable wheel mount. As a result, the turn radius at 90 degrees of steering deflection will be 35.5 mm, and a spin turn would require greater steering deflection.
In one embodiment, the basic speed ratio curves may be functions of the steering angle, as shown in the curves plotted in
In one embodiment, the symmetry of the mowing machine chassis results in the front wheel speed ration curves being mirror images of each other, so only one look up table or fitted function is needed. The third alternative described above can be efficiently implemented because the critical wheel can be identified based on the sign of the steer angle. More specifically, the left front wheel is critical in right turns in which the steer angle is greater than or equal to zero, while the right wheel is critical for left turns. If the critical wheel has a speed command greater than 1.0, all wheel speed commands may be divided by the critical wheel's speed command. If not, the wheel speed commands may be divided by 1.0. In one embodiment, the straight ahead condition may be treated the same as right turns. Both front wheel speed ratios may be 1.0 if the steer angle is zero, so the only special handling needed is including the zero steer angle in the closed interval representing a right turn, and left turns can be represented by an open interval that excludes zero.
In one embodiment of the invention, a 360 degree one piece non-contact rotary position transducer may be used to determine the position of the steered wheel. For example, an RT600 Series sensor from Electro Corp. may be used.
In one embodiment, the kinematically correct wheel speed commands may be calculated for vehicles with independently controlled wheel motors when a suitable steering position sensor is used. The rolling radii for tires mounted on the powered wheels may not be known precisely, and it may be difficult to zero a steering sensor to the exact straight ahead position. In one embodiment, the present invention relates to a methodology which is tolerant of minor variations in sensing the wheel angle, while also providing performance advantages of having the wheel motor speeds coordinated with steering inputs.
In one embodiment, a mechanical system with torque proportioning differentials may be emulated until deteriorating tractive conditions allow one or more wheels to spin out. When the spin out condition is approached, the control logic may limit the torque to the low traction wheels as necessary to maintain the kinematically correct speed ratios for the current steer angle. The torque limitation may be calculated to mimic the performance of a traction control system using automatically actuated wheel brakes, but there is not dissipation involved, merely modifications to the torque commands going to the wheel motors.
In a conventional differential, the output shafts are connected so that they share the torque equally. Such a system is represented in equation form as:
A bevel gear implementation is shown schematically in
In a torque-proportioning differential, the output shafts are connected so that the torque on output one is a fixed multiple of the torque on output two. Such a system is represented in equation form as:
or; equivalently:
Ti=R1Tin
T2=R2Tin
ωin=R1ω1+R2ω2
R1+R2=1
If R1=R2=0.5, the second form represents a conventional differential. A bevel gear implementation of a torque-proportioning differential with a 2:1 torque ratio is shown schematically in
With:
N1=NS1
and
N2=2NS2
giving
R1=⅓
and
R2=⅔
In one embodiment, torque proportioning differential systems with multiple outlets, can be assembled by connecting two or more differentials in series. When this is done the system is represented by these equations:
A three-outlet system supplying ⅓ of the input torque to each outlet is shown in the schematic of
The traditional open differential system has usually been designed to supply each wheel with a torque proportional to its tractive capability. To a first approximation, this results in Ri=Wi/Wtotal.
Viscous damped differentials have been used in many applications. However, the transfer of torque from one outlet to the other is represented by a term of the form: ΔTi=Ci(ωi−ωref). A C value that is too large will impede turning while one that is too low will sacrifice pulling ability in marginal tractive conditions. In one embodiment, the invention matches the value of C or ωref to the operating condition of a mowing vehicle at any moment.
The Ferguson differential, named for its inventor, Harry Ferguson, used an auxiliary gear train and overrunning clutches to limit the ratio of the two output speeds. This effectively allowed the unit to function as an open differential until the limit was reached after which both outputs maintained a fixed ratio to the input. A schematic of a Ferguson differential is shown in
This concept could not be matched to a wide range of maneuvers since setting the limits wide enough to allow short radius turns allowed too large a speed differential when pulling straight ahead. The only successful application was as a center differential in an early all wheel drive automobile. The present invention eliminates this problem by adapting the limiting speed ratio to the vehicle's maneuvers.
Another alternative that has been used in passenger cars is the automated application of the service brakes to limit torque to any wheel that starts to spin out. This is not considered applicable to off road equipment since it dissipates the excess energy scheduled for the low traction wheel instead of recirculating it to help power the other wheels as the Ferguson does.
The hard limits imposed by a Ferguson differential would be difficult to emulate in a stable control system. The following controls are modeled on a traction control system using automatically applied brakes on each wheel. The brake control is based on a hypothetical mechanical system in which a viscous coupling is interposed between the control gear and the overrunning clutch of a Ferguson differential. This produces a rapid but smooth increase in braking torque once the free overspend limit is exceeded, but without the hard discontinuities of the Ferguson differential.
In one embodiment of the invention, two variants of a control scheme may be used. The first, which assumes that the operator inputs a speed command; is shown in block diagram form in
and compute the desired wheel speeds. In speed control mode the next step is to use the desired and actual differential system input speeds to calculate a base torque using any of the generally known closed loop speed control schemes. Base torque for each wheel is then established using the following relations:
T1=R1Tin
T2=R2Tin
. . .
Tn=RnTin
Pseudo-Brake torques are then calculated as follows:
This is the second variant of the control scheme according to one embodiment of the invention, shown in block diagram form in
is then calculated.
The kinematically correct, zero slip, wheel speeds are then calculated, based on the steering configuration and the actual differential system input speed. The next step is to set a base torque using any accelerator pedal position to torque mapping needed to provide the desired control feel. Base torque for each wheel is then established using the following relations:
T1=R1Tin
T2=R2Tin
. . .
Tn=RnTin
Pseudo-Brake torques are then calculated as follows:
In one embodiment, the data that may be used to implement the differential steering and traction control system are shown in the spread sheet labeled as
In
In one embodiment, the chassis controller may implement the basic speed control and determine the speed targets for both motors. The controller will provide the torque setting and speed target to each motor controller. The motor controllers will run in torque control mode unless the target speed is exceeded by n% in which case the torque will be reduced by m% for each additional 1% over speed.
In one embodiment, based on rolling radius measurements of a prototype mowing vehicle, n=about 3% and m=about 15%. This allows driving the vehicle in pure torque control mode until one wheel spins or encounters a severe side slope. However, both of these values may be varied for different vehicles or different performance characteristics.
From the foregoing, it will be observed that numerous variations and modifications may be effected without departing from the spirit and scope of the invention. It is to be understood that no limitation with respect to the specific apparatus illustrated herein is intended or should be inferred. It is, of course, intended to cover by the appended claims all such modifications as fall within the scope of the claims.
This Non-Provisional patent application claims priority based on prior filed U.S. Provisional patent application Ser. No. 60/799,699, filed May 11, 2006, entitled DIFFERENTIAL STEERING AND TRACTION CONTROL FOR ELECTRICALLY PROPELLED MOWER, under 35 USC 119(e).
Number | Date | Country | |
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60799699 | May 2006 | US |