Control systems and architectures for automated wheel loaders

Abstract

Automated wheel loader control system architectures, and corresponding control systems and methods for automated wheel loaders. The control architecture includes an engine speed controller, a vehicle speed controller, a bucket lift speed controller, a bucket tilt speed controller, and a steering controller. The controllers are decoupled from one another. Engine speed, vehicle speed, lift speed, tilt speed and vehicle position signals are used to affect engine throttle, brake pedal, lift valve command, tilt valve command, and steering wheel speed, respectively, to track desired reference signals. The control architecture is formatted such that desired engine speed must be generated, while the other reference signals are inputs to the system. Regardless, the reference profile for each subsystem is generated by corresponding optimization algorithms.

Description

GOVERNMENT SUPPORT CLAUSE

[0001] This invention was made with government support under DE-EE0009200 awarded by the Department of Energy. The government has certain rights in the invention.BACKGROUND

[0002] The present disclosure relates to automated wheel loaders. More particularly, it relates to control systems and related architectures for automated wheel loaders.

[0003] Since the first wheel loader was invented by the Volvo Construction Equipment in 1954, it has been widely-used in different construction and mining sites all over the world. According to the statistical data and predictions from two different institutions, the Off-Highway Research and P&S Intelligence, there exists a growing market for the wheel loader since 2020, which implies that the total population of the wheel loader will reach a higher number. The wheel loader is an intensive energy consuming machine. In the face of the environmental impact and growing demands for wheel loaders, it is necessary to address the challenge of reducing the fuel consumption of the wheel loader without sacrificing or even improving the productivity.

[0004] Human operation of a wheel loader entails coordinated control over a number of actuators / implements, such as the steering wheel, gas / accelerator pedal, brake pedal, transmission, and joystick(s), all of which make the wheel loader work properly. The accelerator pedal controls the engine throttle, while the brake pedal controls the brake torque on the drive shaft. The joystick controls the valve for each of the lift and tilt systems of the wheel loader's bucket (or shovel), which controls the flow to the cylinders. The steering wheel moves a steering cylinder and ultimately controls the steering angle. While experienced operators can capably control these drive and work functions, inefficiencies inevitably arise resulting in excess fuel consumption and / or lower productivity.

[0005] Attempts have been made to develop automated wheel loaders. In general terms, with an automated wheel loader, control algorithms are used to replace the human driver to operate the wheel loader to provide desired vehicle and work tool motion. Given wheel loader complexities, and the strongly nonlinear and highly coupled dynamics, controlling the wheel loader is highly challenging. Currently, human-based calibrations are used in attempt to coordinate the various control algorithms / functions.SUMMARY

[0006] Some aspects of the present disclosure are directed to automated wheel loader control system architectures, and corresponding control systems and methods for automated wheel loaders. In some embodiments, the control architecture includes or provides an engine speed controller, a vehicle speed controller, a bucket lift speed controller, a bucket tilt speed controller, and a steering controller. The controllers are decoupled from one another. Unlike conventional, complicated MIMO (multiple input, multiple output) control strategies, the control architectures of the present disclosure decouple the MIMO arrangement into several SISO (single input, single output) subsystems, which reduce the difficulty to design the corresponding controllers and facilitate implementation in an autonomous wheel loader. Engine speed, vehicle speed, lift speed, tilt speed and vehicle position signals are used to affect engine throttle, brake pedal, lift valve command, tilt valve command, and steering wheel speed, respectively, to track desired reference signals. The control architecture is formatted such that desired engine speed must be generated, while the other reference signals are inputs to the system. Regardless, the reference profile for each subsystem is generated by corresponding optimization algorithms.

[0007] Other aspects of the present disclosure are directed to fully integrated models for automated wheel loaders that implement the control architectures described above.BRIEF DESCRIPTION OF THE DRAWINGS

[0008] FIG. 1 is a schematic diagram of a conventional wheel loader drive system;

[0009] FIG. 2 is a block diagram of an automated wheel loader control architecture or control system in accordance with principles of the present disclosure integrated with the drive system of FIG. 1;

[0010] FIG. 3 is a schematic diagram of an automated wheel loader control system implementing features of the control architecture of FIG. 2;

[0011] FIG. 4 is a block diagram illustrating a method for determining engine speed useful with the control system architectures of the present disclosure;

[0012] FIG. 5 is a schematic diagram of a wheel loader system architecture in accordance with principles of the present disclosure;

[0013] FIG. 6 is a schematic diagram of an example priority valve useful with automated wheel loaders of the present disclosure;

[0014] FIG. 7 is a schematic diagram of a hydraulic working circuit useful with automated wheel loaders of the present disclosure;

[0015] FIG. 8 is a schematic diagram of a hydraulic steering circuit useful with automated wheel loaders of the present disclosure;

[0016] FIG. 9 presents graphs reporting loading cycle external forces described in the Examples section;

[0017] FIG. 10 presents graphs reporting tracking performance of reference speeds described in the Examples section;

[0018] FIG. 11 is a graph reporting tracking performance of vehicle trajectory described in the Examples section;

[0019] FIG. 12 presents graphs reporting simulation torque results described in the Examples section;

[0020] FIG. 13 presents graphs reporting simulation hydraulic pressure and flow results described in the Examples section;

[0021] FIG. 14 is a chart presenting results of energy flow analysis described in the Examples section;

[0022] FIG. 15 presets graphs reporting drivetrain power results described in the Examples section;

[0023] FIG. 16 presets graphs reporting hydraulic circuit power results described in the Examples section; and

[0024] FIG. 17 is a graph reporting percent of energy consumption described in the Examples section.DETAILED DESCRIPTION

[0025] Embodiments of the present disclosure provide automated wheel loader control system architectures, and corresponding control systems and methods for automated wheel loaders. In general terms, control algorithms are utilized to replace the human driver to operate the wheel loader to provide desired vehicle and work tool (e.g., bucket) motions. The control algorithms can assume various forms and incorporate various attributes beneficial to a particular wheel loader design and / or end use environment. In more general terms, the automated wheel loader control system architectures of the present disclosure entail two or more decoupled controllers operating to better ensure both drive function and work function of the wheel loader can be operated in a desired manner, while at the same time ensuring the wheel loader's engine is operated to provide the required power for both functions in real-time. As used throughout the present disclosure, wheel loader “work” or “working” is in reference to movement of the bucket or other working tool of the wheel loader.

[0026] As a point of reference, FIG. 1 represents a typical wheel loader drive system 100. The system 100 includes three major energy-consuming subsystems that require power from the wheel loader's engine 102: a driveline or drivetrain 110, a brake pump 112, and a hydraulic main pump (or main implement pump) 114. The driveline 110 utilizes various components (e.g., torque convertor 120, transmission 122, etc.) to generate desired rotation of wheels 124 via a corresponding wheel axel. The brake pump 112 controls a brake torque on the driveshaft. The hydraulic main pump 114 powers a hydraulic steering circuit 130 and a hydraulic working circuit 132. A priority valve 134 can be provided, and is configured to dictate that fluid (e.g., oil) from the hydraulic main pump 114 to the hydraulic steering circuit 130 is prioritized over fluid being provided by the main pump 114 to the hydraulic working circuit 132. Once the steering circuit 130 is served, the excess supply oil is then available to the working circuit 132.

[0027] The hydraulic steering circuit 130 can include various components conventional to wheel loaders, such as a steering control unit and steering cylinders. The steering control unit can have various forms as are known in the art, and can include a steering valve, a torque generator or gerotor pump, relief valves and check valves. With these and related embodiments, the steering wheel column is connected to a spool of the steering valve. As an operator turns the steering wheel, the steering valve opens to allow flow from the main implement pump 114 to the gerotor pump and then to the steering cylinders. The steering cylinders then turn the wheel loader's wheels. The steering cylinders can be connected (e.g., cross connected) to the steering valve as actuators; a change in the displacement of the cylinders will cause the corresponding change of wheel steering angles. In some configurations, the steering circuit 130 can incorporate a spool valve by means of a rotary spool and a sleeve. Due to the angular position of the spool and sleeve, the flow areas increase or decrease to allow fluid to flow to the gerotor pump and then to the steering cylinders.

[0028] The hydraulic working circuit 132 has, in some embodiments, two functions: lift and tilt. The lift function is for raising and lowering the bucket, with the hydraulic working circuit 132 including corresponding components such as two lift cylinders. The tilt function is for loading and dumping material, with the hydraulic working circuit 132 including corresponding components such as a tilt cylinder. The lift and tilt functions are powered by the main implement pump 114, and each function can be controlled by a proportional valve. Various additional components can also be provided with the working circuit 132 as are known in the art. For example, a check valve for use when the control valve is at a neutral position during transportation for load holding; when the bucket is lowered (lift down or tilt down), the flow required for lift and tilt rod chambers is supplemented by the tank to reduce flow from the main pump 114. This is known as a regeneration function and can be implemented by a regeneration valve.

[0029] Against the above background, an automated wheel loader control architecture or control system 200 in accordance with principles of the present disclosure is shown in FIG. 2 as integrated with the drive system 100 (as described above). A structure or arrangement of the control architecture 200 is separated into two or more decoupled controllers or subsystems. For example, with the embodiment of FIG. 2, the control architecture 200 includes or provides an engine speed controller 210, a vehicle speed controller 212, a lift speed controller 214, and a tilt speed controller 216, with the controllers 210-216 being decoupled from one another. One or more additional controllers can be provided, for example a steering angle controller as described below. As shown, engine speed, vehicle speed, lift speed, tilt speed and vehicle position signals 218 affect engine throttle, brake pedal, lift valve command, tilt valve command, and steering wheel speed, respectively. The control architecture 200 is formatted such that desired engine speed must be generated (referenced generally at 220), while the other reference signals 218 are inputs to the system. Regardless, the reference profile for each subsystem is generated by corresponding optimization algorithms.

[0030] The control architectures of the present disclosure, such as the control architecture 200, uniquely address the constraints presented by wheel loader drive systems, the dynamics of which are otherwise strongly nonlinear and highly coupled. Moreover, a challenge for an automated wheel loader is that both the vehicle drivetrain and the work circuit are driven by the same power source: the wheel loader's internal combustion engine. The control architectures of the present disclosure are configured or formatted to satisfy the needs of both the drivetrain (i.e., vehicle speed) and the work circuit (i.e., lift and tilt speeds). Unlike conventional automated wheel loader control architectures that otherwise utilize a complicated MIMO (multiple input, multiple output) control strategy, the control architectures of the present disclosure decouple the MIMO arrangement into several SISO (single input, single output) subsystems, which reduce the difficulty to design the corresponding controllers and facilitate implementation in an autonomous wheel loader. Each of the subsystem controllers can assume various forms and can be an electronic controller with electrical circuitry configured to process received data as described below. Each controller can include a processor and a memory device, along with other storage devices or suitable components. The processor can be used to execute software and can include one or more microprocessors and / or one or more application specific integrated circuits.

[0031] Control of the wheel loader's engine 102 is vital for operation of every other subsystem because the engine 102 provides power for the entire drive system 100. The reference vehicle speed can be propagated backwards through a drivetrain model to calculate or determine a required engine speed. The required driveshaft speed and torque can be calculated or determined from the vehicle speed and vehicle acceleration, respectively. These are then related to the torque converter turbine speed and torque through the gear ratio, which can be estimated using the shift schedule with an assumed engine speed. The torque converter pump torque and speed, which is engine speed, can then be determined or calculated.

[0032] Without interference from the hydraulic system, tracking this engine speed using throttle can be sufficient to produce the desired vehicle motion. The hydraulic work circuit 132 in particular can demand significant power during certain sections of the cycle. The main pump torque loads the engine as the work circuit moves. In response, the engine controller 210 will increase the throttle and thus increase engine torque to maintain the desired speed. All of the desired vehicle, lift and tilt motions can be tracked or provided with the engine 102 prompted to increase throttle in response to the loads. In many wheel loader applications, the vehicle is pushed to run at the limit of performance in order to move material as fast as possible, which may leave less room for the engine 102 to compensate for unexpected loads. In addition, the work and steering circuits 130, 132 are limited by the available flow from the main pump 114. When the main pump 114 reaches maximum displacement, the flow is then limited by the engine speed. To compensate for these possibilities, in some embodiments the control architecture is programmed or formatted such that the desired engine speed is preemptively increased during times when required power and flow are anticipated to near or exceed the maximum that the engine 102 is providing.

[0033] The control architecture 200 can implement various techniques or controllers for tracking vehicle position with the steering wheel speed control input. The vehicle dynamics features nonlinear coupling between the drivetrain and steering circuit dynamics. Steering control can be segmented into two parts: an outer loop that computes the steering angle necessary to track the vehicle position and an inner (actuator) loop that controls the hydraulic circuit to produce the desired steering angle. The outer loop can, in some embodiments, use a pure pursuit tracking strategy, which determines the steering angle by looking ahead along the reference path and turning towards a future point as is known in the art. The look ahead distance can depend on velocity, and can be tuned to avoid both cutting corners and oscillations. This leaves the actuator controller to track steering angle using the steering wheel speed.

[0034] One example of an automated wheel loader control system 300 implementing features of the control architecture 200 described above is schematically shown in FIG. 3. Commensurate with the above descriptions, five independent or decoupled controllers are provided, each designed to manipulate or control a relevant operational aspect of the wheel loader. For example, an engine speed subsystem 310 includes an engine speed controller (or engine controller) 320 configured and programmed to generate throttle commands to track a desired engine speed by controlling a throttle of the engine. A vehicle speed subsystem 312 includes a vehicle speed controller 322 configured and programmed to generate signals that propel the drivetrain dynamics to reach the desired vehicle speed. A first bucket subsystem 314 includes a lift controller 324 configured and programmed to generate signals dictating a lift position of the vehicle's bucket. A second bucket subsystem 316 includes a tilt controller 326 configured and programmed to generate signals dictating a tilt of the vehicle's bucket. The lift controller 324 and the tilt controller 326 work together to make the bucket (or other implement) follow the desired profile. A steering subsystem 318 includes a steering controller 320 configured and programmed to force the vehicle dynamics to follow the desired profile. While the system 300 has been described as including the first and second bucket subsystems 314, 316, in other embodiments, only a single bucket subsystem (e.g., lift, tilt, etc.) can be provided; in yet other embodiments, additional subsystems (and corresponding controllers) can be included. Regardless, the controllers (e.g., controllers 320-328) are decoupled from one another.

[0035] A desired engine speed (shown in FIG. 3 as an input to the engine speed controller 320) can be determined in various manners. Typically, a duty cycle is provided with the reference profile including desired vehicle speed, vehicle trajectory, lift speed, and tilt speed. The desired engine speed can be obtained from the reference profile, desirably utilizing a strategy that considers the needs of both the drivetrain and the hydraulic circuit.

[0036] One non-limiting example of methods for determining the desired engine speed is represented in FIG. 4. On an upper branch 350 of FIG. 4, the required driveshaft speed and drive torque can be determined by the reference vehicle speed and resistance forces through the driveshaft dynamics, examples of which are provided below. On a lower branch 352 of FIG. 4, the engine speed required by the hydraulic system can be estimated from the cylinder speeds (lift cylinder speed, tilt cylinder speed and steering cylinder speed) and their corresponding areas, assuming the pump is at full displacement. Then, the desired engine speed can take the maximum value of both branches 350, 352 at each time instance in the cycle. This example approach implies that the engine can provide enough power for both systems.

[0037] The control architectures of the present disclosure can be implemented by or utilized with automated wheel loaders, with the independent controllers manipulating relevant parts of the wheel loader. In other examples, the control architectures of the present disclosure can be utilized with fully integrated models for automated wheel loaders as described below.System Overview

[0038] One example of a wheel loader system architecture 400 is shown in FIG. 5. The system 400 is akin to the system 100 (FIG. 1) described above and consists of major energy-consuming subsystems that require power from the wheel loader's engine: a driveline or drivetrain 410, a brake pump 412, and a hydraulic main pump (or main implement pump) 414 that powers a hydraulic working circuit 420 and a hydraulic working circuit 422. A priority valve 424 can be provided, and is configured to dictate that fluid (e.g., oil) from the hydraulic main pump 414 to the hydraulic steering circuit 420 is prioritized over fluid being provided by the main pump 414 to the hydraulic working circuit 422. Once the steering circuit 420 is served, the excess supply oil is then available to the working circuit 422. The engine shaft dynamics are given by Equation 1:Je⁢ω˙e+be⁢ωe=Te-Ttcp-Tmp-TbpEq. (1)where Je is the lumped engine shaft inertia, we is the engine speed, be is the engine friction coefficient, and Te, TtCp, Tmp, Tbp are the engine torque, torque converter pump torque, main hydraulic pump torque, and brake pump torque, respectively. Based on field test data, the brake pump torque can be approximately constant.Drivetrain ModelThe drivetrain model can be based upon a wheel loader using an automatic 4-speed transmission. The torque converter does not have a lock-up clutch and can modeled using a table lookup with the speed ratio, torque ratio, and the K-factor as defined by the following equations:speed⁢ ratio=ωtctωtcpEq. (2)torque⁢ ratio=TtctTtcpEq. (3)K-factor=ωtcpTtcpEq. (4)where ωtc<sub2>p < / sub2>is the turbine (output) speed, ωtc<sub2>p< / sub2>=ωe is the pump (input) speed, Ttc<sub2>t < / sub2>is the turbine (output) torque. At a given speed ratio, the torque ratio and K-factor can be determined. As part of the drivetrain 410, a torque converter turbine connects to a transmission box.Gearshifts are determined by a shift schedule. Since the shift dynamics are fast relative to the rest of the drivetrain dynamics and shift transience has negligible effect on energy consumption, the gear shifts are approximated by a first order delay with a time constant of 0.5 seconds. The dynamics at the final drive shaft are given by:Jf⁢ω˙f=Td-Tf-TbEq. (5)Td=γ⁢TtctEq. (6)where Jf is the lumped inertia of the final drive shaft, ωf is the final drive speed, Td, Tf, Tb are the drive torque, final drive load torque, and brake torque, and γ is the gear ratio.The final drive load torque can be determined by:Tf=Rωkf⁢(μf⁢mg+Fr)Eq. (7)where Rw is the wheel radius, kf is the final drive ratio, μf is the rolling friction coefficient of the wheels, m is the total vehicle mass, g is the acceleration due to gravity, and Fr is the resistance force applied when the bucket interacts with the environment. Due to the low vehicle speeds, the aerodynamic losses can be neglected. The vehicle speed can then be calculated by:v=Rwkf⁢wfEq. (8)Hydraulic System ModelThe hydraulic system can consist of variable displacement pressure and flow compensated pumps, a priority valve, a loader control valve, lift and tilt hydraulic cylinders, steering control unit and steering cylinders.The main implement pump 414 in the wheel loader has variable displacement and is pressure and flow compensated.The torque Tmp and the outlet flow Qmp of the main implement pump 414 are given by:Tmp=xmp⁢Dmp⁢Pmp2⁢πηmpmEq. (9)Qmp=xmp⁢Dmp⁢ωe⁢ηmpv2⁢πEq. (10)where xmp is the displacement fraction of the main pump, Dmp is the full displacement of the main pump, Pmp is main pump outlet pressure, ηmp<sub2>m < / sub2>is the mechanical efficiency of the main pump, ηmp<sub2>v < / sub2>is the volumetric efficiency of the main pump.The main implement pump 414 provides the oil flow required by the working circuit 420 and the steering circuit 422. Hence, the main implement pump 414 outlet pressure can be determined as:dPmpdt=βvmp⁢(Qmp-QEF-QDF)Eq. (11)where Vmp is the main pump chamber control volume, QEF is the working circuit flow, and QCF is the steering circuit flow.The flow regulator in the main pump determines the pump displacement fraction based on the load sensing pressure and main pump outlet pressure.With respect to the priority valve 424, steering is more critical to the operation of the wheel loader, so the priority valve 424 is used to provide oil to the steering circuit 422 first. Once the steering circuit 422 is served, the excess supply oil is then available to the working hydraulic system 420.A schematic of an example of the priority valve 424 is shown in FIG. 6. It is a pilot assisted spool-spring system. A port identified at “P” is from the main pump 414 (FIG. 5). A control flow (“CF”) port connects to the steering circuit 422 (FIG. 5), while an excess flow (“EF”) port connects to the working circuit 420 (FIG. 5). A load sensing input (“LS”) receives feedback from the steering circuit 422. The flow distribution to CF or EF ports is determined by the position of the spool. When there is no steering, due to low LS pressure, the pilot pressure moves the spool to the EF port side, and the flow directs to the EF port. When the steering circuit works, the LS pressure together with the spring overcomes the pilot pressure, which moves the spool inside the valve. Consequently, the flow will be directed to the CF port until the steering circuit meets its requirement. Meanwhile, the flow to the EF port will slowly reduce or even stop. The dynamics of the spool in the priority valve 424 is governed by:my⁢x¨=(PCF-PLS)⁢Ay-by⁢x.-ky(x0+x)-Fnl(x)Eq. (12)where my is the mass of valve spool, x is the displacement of the spool, PCF is the CF port pressure, PLS is the load sensing pressure from the steering circuit, Ay is the effective area of the valve spool, by is the viscous friction coefficient, ky is the valve spring stiffness, x0 is the spring pre-compression value, and Fnl is the mechanical stop force.Returning to FIG. 5, the hydraulic working circuit 420 has two functions: lift and tilt. A schematic example of the hydraulic working circuit 420 is shown in FIG. 7. The lift function is for raising and lowering the bucket via two lift cylinders 430 (one of which is represented in FIG. 7 for simplicity) controlled by a proportional valve 432. The tilt function is for loading and dumping material by tilting the bucket via a tilt cylinder 434 that is controlled by a proportional valve 436.During the bucket lift, the piston chamber pressure of the lift cylinder 430 can be determined as:dP1⁢ldt=βv1⁢l⁢(Cd⁢π⁢dl⁢xvl⁢2ρ⁢(PEF-P1⁢l)-x.l⁢A1⁢l)Eq. (13)where P2l is piston chamber pressure of the lift cylinder 430, V1l is the piston chamber control volume of the lift cylinder 430, dl is the lift valve spool diameter, xvl is the lift valve spool displacement, PEF is the EF (excess flow) port pressure, xl is the lift piston displacement, A1l is the piston chamber area of the lift cylinder 430.The rod chamber pressure of the lift cylinder 430 can be determined as:dP2⁢ldt=βv2⁢l⁢(x.l⁢A2⁢l-Cd⁢π⁢dl⁢xvl⁢2ρ⁢P2⁢l)Eq. (14)where P2l is the rod chamber pressure of the lift cylinder 430, V2l is the rod chamber control volume of the lift cylinder 430, and A2l is the rod chamber area of the lift cylinder 430.The bucket is lowered with the help of the bucket mass, and the flow required for the lift rod chamber is supplemented by the tank, not all from the main implement pump 414 (FIG. 5). During lowering, the piston chamber pressure of the lift cylinder 430 can be determined as:dP1⁢ldt=βv1⁢l⁢(-x.l⁢A1⁢l-Cd⁢π⁢dl⁢xvl⁢2ρ⁢(P1⁢l-PRVs))Eq. (15)where PVR<sub2>s < / sub2>is the regeneration valve setting pressure.The rod chamber pressure of the lift cylinder 430 during lowering can be determined as:dP2⁢ldt=βv2⁢l⁢(Cd⁢π⁢dl⁢xvl⁢2ρ⁢(PRVs-P2⁢l)+x.l⁢A2⁢l)Eq. (16)The lift piston dynamics can be determined as:ml⁢x¨l=P1⁢l⁢A1⁢l-P2⁢l⁢A2⁢l-Fl-ft-bl⁢x.lEq. (17)where ml is mass of the lift piston, Fl is the load force of the lift piston, fl is the frictional force of the lift piston, bris the viscous coefficient of the lift piston.Tilt function can be described in a similar fashion with Eqs. (13)-(17).A schematic example of the hydraulic steering circuit 422 is shown in FIG. 8. The steering control unit consists of a steering valve 450, a torque generator or gerotor pump 452, relief valves, and check valves. The steering wheel column 454 is connected to a spool of the steering valve 450. As the operator turns the steering wheel, the steering valve 450 opens to allow flow from the main implement pump 414 (FIG. 5) to the gerotor pump 452 and then to steering cylinders 456, 458. The steering cylinders 456, 458 turn the wheels.The key part of the hydraulic steering circuit 422 is the spool valve realized by means of a rotary spool and a sleeve. Due to the angular position of the spool and sleeve, the flow areas increase or decrease to allow fluid to flow to the gerotor pump 452 and then to the steering cylinders 456, 458. The two components are the fundamental elements of the hydraulic steering circuit 422 and affect the behavior of the steering and the feeling on the steering wheel controlled by the operator.For the spool dynamics, the spool angle can be determined by integrating the steering wheel speed input as:δ.=usEq. (18)where δ is the spool rotational angle, us is the steering wheel speed input.The gerotor torque, the damping friction torque, and the spring torque drive the sleeve and can be determined from:Jsg⁢θ¨=Tge+bss(δ.-θ.)+kss(δ-θ)Eq. (19)where Jsg is the inertia of the sleeve and the gerotor pump 452 (connected to sleeve), Tge is the torque generated by the gerotor pump 452, bss is the damping coefficient of spool-sleeve, kss is the spring coefficient of spool-sleeve, and θ is the sleeve rotational angle.The gerotor pump 452 ports are connected to the steering valve 450 by means of internal cavities. The gerotor pump 452 in this model is a fixed displacement pump. As the operator turns the steering wheel, the spool rotates first. The relative angle between the spool and the sleeve increases, thus raising the pressure difference across the gerotor pump 452. Torque generated by this pressure difference will drive the sleeve to catch up with the spool.When the spool stays at the left side (relative to FIG. 8), the flow path is as follows: the oil is supplied from port P which is connected to the priority valve CF port to LG port, then goes through the gerotor pump 452. The gerotor pump 452 meters the flow, and the oil goes to port R, then to the right chambers of the steering cylinders 456, 458. The flow from left chambers of the steering cylinders 456, 458 goes towards to the tank. In this case, the gerotor left side pressure can be determined by:dPLGdt=βVLG⁢(Cd⁢AP-LG⁢2ρ⁢(PCF-PLG)-Qge)Eq. (20)where PLG is the gerotor pump left side pressure, VLG is the left side of the gerotor pump control volume, AP-LG is rotary valve open area from port P to port LG, PCF is the control flow pressure, and Qge is the gerotor pump flow.The gerotor right side pressure can be determined as:dPRGdt=βVRG⁢(Qge-Cd⁢ARG-R⁢2ρ⁢(PRG-PR))Eq. (21)where PRG is the gerotor pump right side pressure, VRG is the right side of the gerotor pump control volume, ARG-R is rotary valve open area from port RG to port R, and PR is the right chamber pressure of the steering cylinders 456, 458.The right chamber pressure of the steering cylinders 456, 458 can be determined as:dPRdt=βVR⁢(Cd⁢ARG-R⁢2ρ⁢(PRG-PR)-x.s⁢As)Eq. (22)where Vg is the right side of the steering chamber control volume, xs is the steering piston displacement, and As is the piston area of the steering cylinders 456, 458 since they are symmetric.The left chamber pressure of the steering cylinders 456, 458 can be determined as:dPLdt=βVL⁢(x.⁢As-Cd⁢AL-T⁢2ρ⁢PL)Eq. (23)where PL is left chamber pressure of the steering cylinders 456, 458, VL is the left side of the steering chamber control volume, and AL-T is rotary valve open area from port L to port T.The rotary valve open areas for different ports are a function of the relative angle between spool and sleeve. The case that the spool stays at right side is similar to the description above.The torque Tge and the outlet flow Qge of the gerotor pump 452 can be determined as:Tge=Dge(PLG-PRG)2⁢π⁢ηgmEq. (24)Qge=Dge⁢θ.⁢ηgv2⁢π Eq. (25)Where Dge is the gerotor pump displacement, ηgm is the mechanical efficiency of the gerotor pump 452, and ηgv is the volumetric efficiency of the gerotor pump 452.The steering cylinders 456, 458 are connected to the steering valve as actuators. Two cylinders are cross connected. Change in the displacement of the cylinders 456, 458 will cause the corresponding change of wheel steering angles. The steering piston dynamics can be determined as:ms⁢x¨s=PR⁢AS-PL⁢AS-fs-bs⁢x.sEq. (26)where ms is mass of the steering pistons, fs is the frictional force of the steering pistons, and bs is the viscous coefficient of the steering pistons.The wheel steering angle is proportional to steering piston position from neutral position, and can be determined as:α=KxsEq. (27)where α is the wheel steering angle, and K is the gain from the steering piston position to the wheel steering angle.Vehicle ModelReturning to FIG. 5, the vehicle model combines the outputs of the drivetrain and the steering circuit to calculate the vehicle position and orientation. The vehicle's orientation in a 2-D plane is defined by the heading angle ψ, which can be determined as:ψ=2L⁢v⁢ tan⁢ α2Eq. (28)where L is the wheelbase length. The vehicle 2-D position dynamics can then be determined as:X.=v⁢ cos⁢ψEq. (29)Y.=v⁢ sin⁢ψEq. (30)Model ValidationThe models above can be validated using the control architecture automated wheel loader control system 300 of FIG. 3 (that otherwise reflect the control architecture 200 of FIG. 2). In this regard, and as described above, a key variable to be determined is the engine speed. The strategy of FIG. 4 can be applied. On the upper branch of FIG. 4, the required driveshaft speed ωf and drive torque Td can be determined by the reference vehicle speed and resistance forces through the driveshaft dynamics of Equations (5)-(8). Once the desired vehicle speed v is determined, the driveshaft speed ωf can be determined by:ωf=vkf / RwEq. (31)The final drive load torque Tf can be computed with Equation (7) since all parameters are known. With Equation (5), the drive torque Td can be determined by:Td=Jf⁢ω.f+Tf+TbEq. (32)With this approach, the brake torque Tb is assumed to be 0. Based on the shift schedule, the gear ratio γ can be determined. The turbine torque Ttc<sub2>t < / sub2>can be determined by:Ttct=Td / γEq. (33)The turbine speed can be determined by:ωtct=γωfEq. (34)The torque converter pump speed ωtc<sub2>p < / sub2>which is the same as the engine speed we, can then be determined using the torque converter model of Equations (2)-(4).EXAMPLESEmbodiments and advantages of features of the present disclosure are further illustrated in the following non-limiting examples. The operating conditions and details recited in these examples should not be construed to unduly limit the scope of the present disclosure.System validation was performed using field test data at the subsystem level and at the overall system level. For engine and drivetrain models, engine shaft dynamics and drive shaft dynamics were validated. For the hydraulic working circuit and the hydraulic steering circuit, the flows produced by the model were compared with experimental data with high accuracy. For this fully integrated system model, the difference between the measured fuel consumption of the loading cycle and the simulation result is 2.3%.The fully integrated and validated model could then be simulated for a typical loading cycle. A typical loading cycle includes six phases:Phase 1—Digging. The vehicle is at the material pile. The lift piston extends to lift the bucket and the tilt piston extends to load the material while the vehicle drives forward slowly.Phase 2—Transport Reverse 1. The vehicle drive backwards away from the pile and turns towards the truck.Phase 3—Transport Forward 1. The wheel loader drives forward to the truck while the lift piston extends to raise the full bucket.Phase 4—Dumping. The tilt piston retracts to dump the material into the truck.

[0089] Phase 5—Transport Reverse 2. The wheel loader drives backward to the starting point while the lift piston retracts to lower the empty bucket.

[0090] Phase 6—Transport Forward 2. The vehicle drives forward towards the pile of material.

[0091] The load forces during the loading cycle for lift and tilt cylinders are necessary for simulation. They can be determined as follows. Lift and tilt cylinder motion can be transformed to bucket motion and orientation through the linkage mechanism, and vice-versa. The interaction force between the bucket and the soil can be determined by the fundamental earthmoving equation during digging as described in Lee et al., “Force-balancing algorithm to remove the discontinuity in soil force during wheel loader excavation.”Journal of Mechanical Science and Technology, 32(10), pp. 4951-4957 (2018), the entire teachings of which are incorporated herein by reference. During the transportation before dumping, the constant weight of the material and bucket is the dominant force. The material weight decreases parabolically during dumping. The constant weight of the empty bucket is the dominant force for the rest of the transportation.

[0092] These bucket forces can then be converted to lift and tilt cylinder forces through the linkage mechanism. The force on the bucket acting against vehicle motion during digging is also the resistance force. The loading cycle external forces are shown in FIG. 9.

[0093] Simulation was conducted via MATLAB / Simulink. The tracking performance of the system compared to the references during Phases 1-6 is shown in FIG. 10 (as shown, the “Desired” plot line and the “Simulated” plot line are nearly identical). The Root-Mean-Squared-Error (RMSE) for the lift, tilt, vehicle, and engine speed tracking were found to be 3.1%, 0.8%, 4.9% and 4.3%, which indicates the control architecture is effective.

[0094] The tracking performance of the vehicle trajectory is shown in FIG. 11. The vehicle starts at point 1. From point 1 to point 2, the vehicle fills the bucket at the Pile. After that, the vehicle reverses to point 3 and moves forward to the Truck. At point 4, the vehicle dumps the material. The plot lines of FIG. 11 reflect that there is tracking overshoot during steep steering, but it settles down eventually. After dumping the material, the vehicle reverses from point 4 to point 5 and then advances to the Pile for the next cycle.

[0095] The torques acting on the engine shaft are shown in FIG. 12. The main pump torque is large during digging and bucket lifting, when the working circuit is doing work. The torque converter torque is never small and dominates the engine load during the transportation phases to accelerate the vehicle. It is also the highest during digging, since the resistance force induces a large load on the vehicle as it moves into the pile.

[0096] The pressure and flow for the hydraulic circuit is shown in FIG. 13. The main pump pressure is always larger than the highest working or steering circuit pressures unless the working circuit is load holding or lowering. During these cases, the pressure can be held constant or lowering without power from the pump.

[0097] The flow required by the working circuit is mainly during bucket filling and lifting, when there are cylinder motions. There is minimal flow to the working circuit during bucket lowering, when it uses regeneration. The flow of the steering circuit is relatively lower than the flow of the working circuit. The main pump provides the total flow to both circuits.

[0098] Energy flow analysis was performed based on the simulation results. The subsystem energy flow is shown in FIG. 14.

[0099] The first part of the energy flow from the engine is to the drivetrain. Specifically, the flow of energy in the drivetrain is from the engine to the torque converter, then through the transmission, and finally through the final drive shaft to the wheels. The torque converter takes 65.7% of the energy produced by the engine. 50.8% of the energy is transferred to the wheels to produce vehicle motion. The detailed drivetrain power analysis results are shown in FIG. 15. The torque converter gets most engine power during vehicle acceleration. The torque converter output follows a similar trend to the drivetrain input power. The torque converter loss is highest during digging. The reason is as follows. The vehicle speed is low at that time while the working circuit requires high pump flow, which means the engine speed is high. The ratio of the torque converter pump speed and turbine speed is high, so the torque converter operates at low efficiency.

[0100] Returning to FIG. 14, the second portion of the energy flow from the engine goes to the hydraulic circuit. Specifically, 22.7% of the energy from the engine flows to the main pump. The priority valve distributes the main pump output energy between the loader valve and the steering valve, with some throttling losses. Finally, energy flows into the lift and tilt cylinders to generate work tool motion, while the steering cylinders receive energy from the steering valve to turn the wheels. The detailed hydraulic circuit power analysis results are shown in FIG. 16. The energy loss of a subsystem is the difference between the input energy and the output energy. Main pump losses are mainly caused by low displacement operation or idling. The lifting loss is mainly the valve throttling loss during bucket lifting. The tilt loss during bucket filling is greater than the lift loss because the main pump provides high pressure due to the greater load force on the lift cylinder. The pressure drop from the main pump pressure to the tilt pressure is higher, which results in more throttling losses for the tilt valve. Steering losses are approximately half of the input energy due to throttling losses in valves and gerotor pump.

[0101] FIG. 17 shows the energy distribution across the entire cycle for the subsystems. Drivetrain transmits most energy from the engine. The energy consumption of the brake pump is approximately the same across the cycle. However, the relative energy consumption of the drivetrain and main pump is variable throughout the cycle. During transportation, especially during acceleration, the drivetrain accounts for a large portion of the engine output energy. While the main pump increases its share of energy consumption during bucket lifting. In addition, during digging, the main pump accounts for a larger portion of the engine output energy than any other time. It's worth noting that the drivetrain still consumes almost half of the engine's output energy during digging.

[0102] The systems and methods of the present disclosure provide a marked improvement over previous designs. The control system architectures of the present disclosure ensure both the drive function and the work function of an automated wheel loader can be operated in a desired manner while the engine can be operated to provide the required power for both functions in real-time. For example, the control architectures of the present disclosure decouple the MIMO system into several SISO systems, leading to the facilitation of designing controllers and implementing them. In other aspects, a fully integrated wheel loader model is provided, which includes engine, drivetrain, working circuit, steering circuit, and vehicle. Based on the model and the control architecture, a tracking problem is formulated with a typical loading cycle. The 2.3% difference between the measured and simulated fuel consumption verifies the accuracy of the model and the effectiveness of the control architecture. The control architectures and models of the present disclosure can facilitate the model-based control design and performance optimization for an autonomous wheel loader.

[0103] Although the present disclosure has been described with reference to preferred embodiments, workers skilled in the art will recognize that changes can be made in form and detail without departing from the spirit and scope of the present disclosure. For example, the control system architectures are useful not only for automated wheel loaders, but also other off-road vehicles.

Examples

examples

Embodiments and advantages of features of the present disclosure are further illustrated in the following non-limiting examples. The operating conditions and details recited in these examples should not be construed to unduly limit the scope of the present disclosure.

System validation was performed using field test data at the subsystem level and at the overall system level. For engine and drivetrain models, engine shaft dynamics and drive shaft dynamics were validated. For the hydraulic working circuit and the hydraulic steering circuit, the flows produced by the model were compared with experimental data with high accuracy. For this fully integrated system model, the difference between the measured fuel consumption of the loading cycle and the simulation result is 2.3%.

The fully integrated and validated model could then be simulated for a typical loading cycle. A typical loading cycle includes six phases:Phase 1—Digging. The vehicle is at the material pile. The lift piston extends...

GOVERNMENT SUPPORT CLAUSE

[0001] This invention was made with government support under DE-EE0009200 awarded by the Department of Energy. The government has certain rights in the invention.BACKGROUND

[0002] The present disclosure relates to automated wheel loaders. More particularly, it relates to control systems and related architectures for automated wheel loaders.

[0003] Since the first wheel loader was invented by the Volvo Construction Equipment in 1954, it has been widely-used in different construction and mining sites all over the world. According to the statistical data and predictions from two different institutions, the Off-Highway Research and P&S Intelligence, there exists a growing market for the wheel loader since 2020, which implies that the total population of the wheel loader will reach a higher number. The wheel loader is an intensive energy consuming machine. In the face of the environmental impact and growing demands for wheel loaders, it is necessary to address the challenge of reducing the fuel consumption of the wheel loader without sacrificing or even improving the productivity.

[0004] Human operation of a wheel loader entails coordinated control over a number of actuators / implements, such as the steering wheel, gas / accelerator pedal, brake pedal, transmission, and joystick(s), all of which make the wheel loader work properly. The accelerator pedal controls the engine throttle, while the brake pedal controls the brake torque on the drive shaft. The joystick controls the valve for each of the lift and tilt systems of the wheel loader's bucket (or shovel), which controls the flow to the cylinders. The steering wheel moves a steering cylinder and ultimately controls the steering angle. While experienced operators can capably control these drive and work functions, inefficiencies inevitably arise resulting in excess fuel consumption and / or lower productivity.

[0005] Attempts have been made to develop automated wheel loaders. In general terms, with an automated wheel loader, control algorithms are used to replace the human driver to operate the wheel loader to provide desired vehicle and work tool motion. Given wheel loader complexities, and the strongly nonlinear and highly coupled dynamics, controlling the wheel loader is highly challenging. Currently, human-based calibrations are used in attempt to coordinate the various control algorithms / functions.SUMMARY

[0006] Some aspects of the present disclosure are directed to automated wheel loader control system architectures, and corresponding control systems and methods for automated wheel loaders. In some embodiments, the control architecture includes or provides an engine speed controller, a vehicle speed controller, a bucket lift speed controller, a bucket tilt speed controller, and a steering controller. The controllers are decoupled from one another. Unlike conventional, complicated MIMO (multiple input, multiple output) control strategies, the control architectures of the present disclosure decouple the MIMO arrangement into several SISO (single input, single output) subsystems, which reduce the difficulty to design the corresponding controllers and facilitate implementation in an autonomous wheel loader. Engine speed, vehicle speed, lift speed, tilt speed and vehicle position signals are used to affect engine throttle, brake pedal, lift valve command, tilt valve command, and steering wheel speed, respectively, to track desired reference signals. The control architecture is formatted such that desired engine speed must be generated, while the other reference signals are inputs to the system. Regardless, the reference profile for each subsystem is generated by corresponding optimization algorithms.

[0007] Other aspects of the present disclosure are directed to fully integrated models for automated wheel loaders that implement the control architectures described above.BRIEF DESCRIPTION OF THE DRAWINGS

[0008] FIG. 1 is a schematic diagram of a conventional wheel loader drive system;

[0009] FIG. 2 is a block diagram of an automated wheel loader control architecture or control system in accordance with principles of the present disclosure integrated with the drive system of FIG. 1;

[0010] FIG. 3 is a schematic diagram of an automated wheel loader control system implementing features of the control architecture of FIG. 2;

[0011] FIG. 4 is a block diagram illustrating a method for determining engine speed useful with the control system architectures of the present disclosure;

[0012] FIG. 5 is a schematic diagram of a wheel loader system architecture in accordance with principles of the present disclosure;

[0013] FIG. 6 is a schematic diagram of an example priority valve useful with automated wheel loaders of the present disclosure;

[0014] FIG. 7 is a schematic diagram of a hydraulic working circuit useful with automated wheel loaders of the present disclosure;

[0015] FIG. 8 is a schematic diagram of a hydraulic steering circuit useful with automated wheel loaders of the present disclosure;

[0016] FIG. 9 presents graphs reporting loading cycle external forces described in the Examples section;

[0017] FIG. 10 presents graphs reporting tracking performance of reference speeds described in the Examples section;

[0018] FIG. 11 is a graph reporting tracking performance of vehicle trajectory described in the Examples section;

[0019] FIG. 12 presents graphs reporting simulation torque results described in the Examples section;

[0020] FIG. 13 presents graphs reporting simulation hydraulic pressure and flow results described in the Examples section;

[0021] FIG. 14 is a chart presenting results of energy flow analysis described in the Examples section;

[0022] FIG. 15 presets graphs reporting drivetrain power results described in the Examples section;

[0023] FIG. 16 presets graphs reporting hydraulic circuit power results described in the Examples section; and

[0024] FIG. 17 is a graph reporting percent of energy consumption described in the Examples section.DETAILED DESCRIPTION

[0025] Embodiments of the present disclosure provide automated wheel loader control system architectures, and corresponding control systems and methods for automated wheel loaders. In general terms, control algorithms are utilized to replace the human driver to operate the wheel loader to provide desired vehicle and work tool (e.g., bucket) motions. The control algorithms can assume various forms and incorporate various attributes beneficial to a particular wheel loader design and / or end use environment. In more general terms, the automated wheel loader control system architectures of the present disclosure entail two or more decoupled controllers operating to better ensure both drive function and work function of the wheel loader can be operated in a desired manner, while at the same time ensuring the wheel loader's engine is operated to provide the required power for both functions in real-time. As used throughout the present disclosure, wheel loader “work” or “working” is in reference to movement of the bucket or other working tool of the wheel loader.

[0026] As a point of reference, FIG. 1 represents a typical wheel loader drive system 100. The system 100 includes three major energy-consuming subsystems that require power from the wheel loader's engine 102: a driveline or drivetrain 110, a brake pump 112, and a hydraulic main pump (or main implement pump) 114. The driveline 110 utilizes various components (e.g., torque convertor 120, transmission 122, etc.) to generate desired rotation of wheels 124 via a corresponding wheel axel. The brake pump 112 controls a brake torque on the driveshaft. The hydraulic main pump 114 powers a hydraulic steering circuit 130 and a hydraulic working circuit 132. A priority valve 134 can be provided, and is configured to dictate that fluid (e.g., oil) from the hydraulic main pump 114 to the hydraulic steering circuit 130 is prioritized over fluid being provided by the main pump 114 to the hydraulic working circuit 132. Once the steering circuit 130 is served, the excess supply oil is then available to the working circuit 132.

[0027] The hydraulic steering circuit 130 can include various components conventional to wheel loaders, such as a steering control unit and steering cylinders. The steering control unit can have various forms as are known in the art, and can include a steering valve, a torque generator or gerotor pump, relief valves and check valves. With these and related embodiments, the steering wheel column is connected to a spool of the steering valve. As an operator turns the steering wheel, the steering valve opens to allow flow from the main implement pump 114 to the gerotor pump and then to the steering cylinders. The steering cylinders then turn the wheel loader's wheels. The steering cylinders can be connected (e.g., cross connected) to the steering valve as actuators; a change in the displacement of the cylinders will cause the corresponding change of wheel steering angles. In some configurations, the steering circuit 130 can incorporate a spool valve by means of a rotary spool and a sleeve. Due to the angular position of the spool and sleeve, the flow areas increase or decrease to allow fluid to flow to the gerotor pump and then to the steering cylinders.

[0028] The hydraulic working circuit 132 has, in some embodiments, two functions: lift and tilt. The lift function is for raising and lowering the bucket, with the hydraulic working circuit 132 including corresponding components such as two lift cylinders. The tilt function is for loading and dumping material, with the hydraulic working circuit 132 including corresponding components such as a tilt cylinder. The lift and tilt functions are powered by the main implement pump 114, and each function can be controlled by a proportional valve. Various additional components can also be provided with the working circuit 132 as are known in the art. For example, a check valve for use when the control valve is at a neutral position during transportation for load holding; when the bucket is lowered (lift down or tilt down), the flow required for lift and tilt rod chambers is supplemented by the tank to reduce flow from the main pump 114. This is known as a regeneration function and can be implemented by a regeneration valve.

[0029] Against the above background, an automated wheel loader control architecture or control system 200 in accordance with principles of the present disclosure is shown in FIG. 2 as integrated with the drive system 100 (as described above). A structure or arrangement of the control architecture 200 is separated into two or more decoupled controllers or subsystems. For example, with the embodiment of FIG. 2, the control architecture 200 includes or provides an engine speed controller 210, a vehicle speed controller 212, a lift speed controller 214, and a tilt speed controller 216, with the controllers 210-216 being decoupled from one another. One or more additional controllers can be provided, for example a steering angle controller as described below. As shown, engine speed, vehicle speed, lift speed, tilt speed and vehicle position signals 218 affect engine throttle, brake pedal, lift valve command, tilt valve command, and steering wheel speed, respectively. The control architecture 200 is formatted such that desired engine speed must be generated (referenced generally at 220), while the other reference signals 218 are inputs to the system. Regardless, the reference profile for each subsystem is generated by corresponding optimization algorithms.

[0030] The control architectures of the present disclosure, such as the control architecture 200, uniquely address the constraints presented by wheel loader drive systems, the dynamics of which are otherwise strongly nonlinear and highly coupled. Moreover, a challenge for an automated wheel loader is that both the vehicle drivetrain and the work circuit are driven by the same power source: the wheel loader's internal combustion engine. The control architectures of the present disclosure are configured or formatted to satisfy the needs of both the drivetrain (i.e., vehicle speed) and the work circuit (i.e., lift and tilt speeds). Unlike conventional automated wheel loader control architectures that otherwise utilize a complicated MIMO (multiple input, multiple output) control strategy, the control architectures of the present disclosure decouple the MIMO arrangement into several SISO (single input, single output) subsystems, which reduce the difficulty to design the corresponding controllers and facilitate implementation in an autonomous wheel loader. Each of the subsystem controllers can assume various forms and can be an electronic controller with electrical circuitry configured to process received data as described below. Each controller can include a processor and a memory device, along with other storage devices or suitable components. The processor can be used to execute software and can include one or more microprocessors and / or one or more application specific integrated circuits.

[0031] Control of the wheel loader's engine 102 is vital for operation of every other subsystem because the engine 102 provides power for the entire drive system 100. The reference vehicle speed can be propagated backwards through a drivetrain model to calculate or determine a required engine speed. The required driveshaft speed and torque can be calculated or determined from the vehicle speed and vehicle acceleration, respectively. These are then related to the torque converter turbine speed and torque through the gear ratio, which can be estimated using the shift schedule with an assumed engine speed. The torque converter pump torque and speed, which is engine speed, can then be determined or calculated.

[0032] Without interference from the hydraulic system, tracking this engine speed using throttle can be sufficient to produce the desired vehicle motion. The hydraulic work circuit 132 in particular can demand significant power during certain sections of the cycle. The main pump torque loads the engine as the work circuit moves. In response, the engine controller 210 will increase the throttle and thus increase engine torque to maintain the desired speed. All of the desired vehicle, lift and tilt motions can be tracked or provided with the engine 102 prompted to increase throttle in response to the loads. In many wheel loader applications, the vehicle is pushed to run at the limit of performance in order to move material as fast as possible, which may leave less room for the engine 102 to compensate for unexpected loads. In addition, the work and steering circuits 130, 132 are limited by the available flow from the main pump 114. When the main pump 114 reaches maximum displacement, the flow is then limited by the engine speed. To compensate for these possibilities, in some embodiments the control architecture is programmed or formatted such that the desired engine speed is preemptively increased during times when required power and flow are anticipated to near or exceed the maximum that the engine 102 is providing.

[0033] The control architecture 200 can implement various techniques or controllers for tracking vehicle position with the steering wheel speed control input. The vehicle dynamics features nonlinear coupling between the drivetrain and steering circuit dynamics. Steering control can be segmented into two parts: an outer loop that computes the steering angle necessary to track the vehicle position and an inner (actuator) loop that controls the hydraulic circuit to produce the desired steering angle. The outer loop can, in some embodiments, use a pure pursuit tracking strategy, which determines the steering angle by looking ahead along the reference path and turning towards a future point as is known in the art. The look ahead distance can depend on velocity, and can be tuned to avoid both cutting corners and oscillations. This leaves the actuator controller to track steering angle using the steering wheel speed.

[0034] One example of an automated wheel loader control system 300 implementing features of the control architecture 200 described above is schematically shown in FIG. 3. Commensurate with the above descriptions, five independent or decoupled controllers are provided, each designed to manipulate or control a relevant operational aspect of the wheel loader. For example, an engine speed subsystem 310 includes an engine speed controller (or engine controller) 320 configured and programmed to generate throttle commands to track a desired engine speed by controlling a throttle of the engine. A vehicle speed subsystem 312 includes a vehicle speed controller 322 configured and programmed to generate signals that propel the drivetrain dynamics to reach the desired vehicle speed. A first bucket subsystem 314 includes a lift controller 324 configured and programmed to generate signals dictating a lift position of the vehicle's bucket. A second bucket subsystem 316 includes a tilt controller 326 configured and programmed to generate signals dictating a tilt of the vehicle's bucket. The lift controller 324 and the tilt controller 326 work together to make the bucket (or other implement) follow the desired profile. A steering subsystem 318 includes a steering controller 320 configured and programmed to force the vehicle dynamics to follow the desired profile. While the system 300 has been described as including the first and second bucket subsystems 314, 316, in other embodiments, only a single bucket subsystem (e.g., lift, tilt, etc.) can be provided; in yet other embodiments, additional subsystems (and corresponding controllers) can be included. Regardless, the controllers (e.g., controllers 320-328) are decoupled from one another.

[0035] A desired engine speed (shown in FIG. 3 as an input to the engine speed controller 320) can be determined in various manners. Typically, a duty cycle is provided with the reference profile including desired vehicle speed, vehicle trajectory, lift speed, and tilt speed. The desired engine speed can be obtained from the reference profile, desirably utilizing a strategy that considers the needs of both the drivetrain and the hydraulic circuit.

[0036] One non-limiting example of methods for determining the desired engine speed is represented in FIG. 4. On an upper branch 350 of FIG. 4, the required driveshaft speed and drive torque can be determined by the reference vehicle speed and resistance forces through the driveshaft dynamics, examples of which are provided below. On a lower branch 352 of FIG. 4, the engine speed required by the hydraulic system can be estimated from the cylinder speeds (lift cylinder speed, tilt cylinder speed and steering cylinder speed) and their corresponding areas, assuming the pump is at full displacement. Then, the desired engine speed can take the maximum value of both branches 350, 352 at each time instance in the cycle. This example approach implies that the engine can provide enough power for both systems.

[0037] The control architectures of the present disclosure can be implemented by or utilized with automated wheel loaders, with the independent controllers manipulating relevant parts of the wheel loader. In other examples, the control architectures of the present disclosure can be utilized with fully integrated models for automated wheel loaders as described below.System Overview

[0038] One example of a wheel loader system architecture 400 is shown in FIG. 5. The system 400 is akin to the system 100 (FIG. 1) described above and consists of major energy-consuming subsystems that require power from the wheel loader's engine: a driveline or drivetrain 410, a brake pump 412, and a hydraulic main pump (or main implement pump) 414 that powers a hydraulic working circuit 420 and a hydraulic working circuit 422. A priority valve 424 can be provided, and is configured to dictate that fluid (e.g., oil) from the hydraulic main pump 414 to the hydraulic steering circuit 420 is prioritized over fluid being provided by the main pump 414 to the hydraulic working circuit 422. Once the steering circuit 420 is served, the excess supply oil is then available to the working circuit 422. The engine shaft dynamics are given by Equation 1:Je⁢ω˙e+be⁢ωe=Te-Ttcp-Tmp-TbpEq. (1)where Je is the lumped engine shaft inertia, we is the engine speed, be is the engine friction coefficient, and Te, TtCp, Tmp, Tbp are the engine torque, torque converter pump torque, main hydraulic pump torque, and brake pump torque, respectively. Based on field test data, the brake pump torque can be approximately constant.Drivetrain ModelThe drivetrain model can be based upon a wheel loader using an automatic 4-speed transmission. The torque converter does not have a lock-up clutch and can modeled using a table lookup with the speed ratio, torque ratio, and the K-factor as defined by the following equations:speed⁢ ratio=ωtctωtcpEq. (2)torque⁢ ratio=TtctTtcpEq. (3)K-factor=ωtcpTtcpEq. (4)where ωtc<sub2>p < / sub2>is the turbine (output) speed, ωtc<sub2>p< / sub2>=ωe is the pump (input) speed, Ttc<sub2>t < / sub2>is the turbine (output) torque. At a given speed ratio, the torque ratio and K-factor can be determined. As part of the drivetrain 410, a torque converter turbine connects to a transmission box.Gearshifts are determined by a shift schedule. Since the shift dynamics are fast relative to the rest of the drivetrain dynamics and shift transience has negligible effect on energy consumption, the gear shifts are approximated by a first order delay with a time constant of 0.5 seconds. The dynamics at the final drive shaft are given by:Jf⁢ω˙f=Td-Tf-TbEq. (5)Td=γ⁢TtctEq. (6)where Jf is the lumped inertia of the final drive shaft, ωf is the final drive speed, Td, Tf, Tb are the drive torque, final drive load torque, and brake torque, and γ is the gear ratio.The final drive load torque can be determined by:Tf=Rωkf⁢(μf⁢mg+Fr)Eq. (7)where Rw is the wheel radius, kf is the final drive ratio, μf is the rolling friction coefficient of the wheels, m is the total vehicle mass, g is the acceleration due to gravity, and Fr is the resistance force applied when the bucket interacts with the environment. Due to the low vehicle speeds, the aerodynamic losses can be neglected. The vehicle speed can then be calculated by:v=Rwkf⁢wfEq. (8)Hydraulic System ModelThe hydraulic system can consist of variable displacement pressure and flow compensated pumps, a priority valve, a loader control valve, lift and tilt hydraulic cylinders, steering control unit and steering cylinders.The main implement pump 414 in the wheel loader has variable displacement and is pressure and flow compensated.The torque Tmp and the outlet flow Qmp of the main implement pump 414 are given by:Tmp=xmp⁢Dmp⁢Pmp2⁢πηmpmEq. (9)Qmp=xmp⁢Dmp⁢ωe⁢ηmpv2⁢πEq. (10)where xmp is the displacement fraction of the main pump, Dmp is the full displacement of the main pump, Pmp is main pump outlet pressure, ηmp<sub2>m < / sub2>is the mechanical efficiency of the main pump, ηmp<sub2>v < / sub2>is the volumetric efficiency of the main pump.The main implement pump 414 provides the oil flow required by the working circuit 420 and the steering circuit 422. Hence, the main implement pump 414 outlet pressure can be determined as:dPmpdt=βvmp⁢(Qmp-QEF-QDF)Eq. (11)where Vmp is the main pump chamber control volume, QEF is the working circuit flow, and QCF is the steering circuit flow.The flow regulator in the main pump determines the pump displacement fraction based on the load sensing pressure and main pump outlet pressure.With respect to the priority valve 424, steering is more critical to the operation of the wheel loader, so the priority valve 424 is used to provide oil to the steering circuit 422 first. Once the steering circuit 422 is served, the excess supply oil is then available to the working hydraulic system 420.A schematic of an example of the priority valve 424 is shown in FIG. 6. It is a pilot assisted spool-spring system. A port identified at “P” is from the main pump 414 (FIG. 5). A control flow (“CF”) port connects to the steering circuit 422 (FIG. 5), while an excess flow (“EF”) port connects to the working circuit 420 (FIG. 5). A load sensing input (“LS”) receives feedback from the steering circuit 422. The flow distribution to CF or EF ports is determined by the position of the spool. When there is no steering, due to low LS pressure, the pilot pressure moves the spool to the EF port side, and the flow directs to the EF port. When the steering circuit works, the LS pressure together with the spring overcomes the pilot pressure, which moves the spool inside the valve. Consequently, the flow will be directed to the CF port until the steering circuit meets its requirement. Meanwhile, the flow to the EF port will slowly reduce or even stop. The dynamics of the spool in the priority valve 424 is governed by:my⁢x¨=(PCF-PLS)⁢Ay-by⁢x.-ky(x0+x)-Fnl(x)Eq. (12)where my is the mass of valve spool, x is the displacement of the spool, PCF is the CF port pressure, PLS is the load sensing pressure from the steering circuit, Ay is the effective area of the valve spool, by is the viscous friction coefficient, ky is the valve spring stiffness, x0 is the spring pre-compression value, and Fnl is the mechanical stop force.Returning to FIG. 5, the hydraulic working circuit 420 has two functions: lift and tilt. A schematic example of the hydraulic working circuit 420 is shown in FIG. 7. The lift function is for raising and lowering the bucket via two lift cylinders 430 (one of which is represented in FIG. 7 for simplicity) controlled by a proportional valve 432. The tilt function is for loading and dumping material by tilting the bucket via a tilt cylinder 434 that is controlled by a proportional valve 436.During the bucket lift, the piston chamber pressure of the lift cylinder 430 can be determined as:dP1⁢ldt=βv1⁢l⁢(Cd⁢π⁢dl⁢xvl⁢2ρ⁢(PEF-P1⁢l)-x.l⁢A1⁢l)Eq. (13)where P2l is piston chamber pressure of the lift cylinder 430, V1l is the piston chamber control volume of the lift cylinder 430, dl is the lift valve spool diameter, xvl is the lift valve spool displacement, PEF is the EF (excess flow) port pressure, xl is the lift piston displacement, A1l is the piston chamber area of the lift cylinder 430.The rod chamber pressure of the lift cylinder 430 can be determined as:dP2⁢ldt=βv2⁢l⁢(x.l⁢A2⁢l-Cd⁢π⁢dl⁢xvl⁢2ρ⁢P2⁢l)Eq. (14)where P2l is the rod chamber pressure of the lift cylinder 430, V2l is the rod chamber control volume of the lift cylinder 430, and A2l is the rod chamber area of the lift cylinder 430.The bucket is lowered with the help of the bucket mass, and the flow required for the lift rod chamber is supplemented by the tank, not all from the main implement pump 414 (FIG. 5). During lowering, the piston chamber pressure of the lift cylinder 430 can be determined as:dP1⁢ldt=βv1⁢l⁢(-x.l⁢A1⁢l-Cd⁢π⁢dl⁢xvl⁢2ρ⁢(P1⁢l-PRVs))Eq. (15)where PVR<sub2>s < / sub2>is the regeneration valve setting pressure.The rod chamber pressure of the lift cylinder 430 during lowering can be determined as:dP2⁢ldt=βv2⁢l⁢(Cd⁢π⁢dl⁢xvl⁢2ρ⁢(PRVs-P2⁢l)+x.l⁢A2⁢l)Eq. (16)The lift piston dynamics can be determined as:ml⁢x¨l=P1⁢l⁢A1⁢l-P2⁢l⁢A2⁢l-Fl-ft-bl⁢x.lEq. (17)where ml is mass of the lift piston, Fl is the load force of the lift piston, fl is the frictional force of the lift piston, bris the viscous coefficient of the lift piston.Tilt function can be described in a similar fashion with Eqs. (13)-(17).A schematic example of the hydraulic steering circuit 422 is shown in FIG. 8. The steering control unit consists of a steering valve 450, a torque generator or gerotor pump 452, relief valves, and check valves. The steering wheel column 454 is connected to a spool of the steering valve 450. As the operator turns the steering wheel, the steering valve 450 opens to allow flow from the main implement pump 414 (FIG. 5) to the gerotor pump 452 and then to steering cylinders 456, 458. The steering cylinders 456, 458 turn the wheels.The key part of the hydraulic steering circuit 422 is the spool valve realized by means of a rotary spool and a sleeve. Due to the angular position of the spool and sleeve, the flow areas increase or decrease to allow fluid to flow to the gerotor pump 452 and then to the steering cylinders 456, 458. The two components are the fundamental elements of the hydraulic steering circuit 422 and affect the behavior of the steering and the feeling on the steering wheel controlled by the operator.For the spool dynamics, the spool angle can be determined by integrating the steering wheel speed input as:δ.=usEq. (18)where δ is the spool rotational angle, us is the steering wheel speed input.The gerotor torque, the damping friction torque, and the spring torque drive the sleeve and can be determined from:Jsg⁢θ¨=Tge+bss(δ.-θ.)+kss(δ-θ)Eq. (19)where Jsg is the inertia of the sleeve and the gerotor pump 452 (connected to sleeve), Tge is the torque generated by the gerotor pump 452, bss is the damping coefficient of spool-sleeve, kss is the spring coefficient of spool-sleeve, and θ is the sleeve rotational angle.The gerotor pump 452 ports are connected to the steering valve 450 by means of internal cavities. The gerotor pump 452 in this model is a fixed displacement pump. As the operator turns the steering wheel, the spool rotates first. The relative angle between the spool and the sleeve increases, thus raising the pressure difference across the gerotor pump 452. Torque generated by this pressure difference will drive the sleeve to catch up with the spool.When the spool stays at the left side (relative to FIG. 8), the flow path is as follows: the oil is supplied from port P which is connected to the priority valve CF port to LG port, then goes through the gerotor pump 452. The gerotor pump 452 meters the flow, and the oil goes to port R, then to the right chambers of the steering cylinders 456, 458. The flow from left chambers of the steering cylinders 456, 458 goes towards to the tank. In this case, the gerotor left side pressure can be determined by:dPLGdt=βVLG⁢(Cd⁢AP-LG⁢2ρ⁢(PCF-PLG)-Qge)Eq. (20)where PLG is the gerotor pump left side pressure, VLG is the left side of the gerotor pump control volume, AP-LG is rotary valve open area from port P to port LG, PCF is the control flow pressure, and Qge is the gerotor pump flow.The gerotor right side pressure can be determined as:dPRGdt=βVRG⁢(Qge-Cd⁢ARG-R⁢2ρ⁢(PRG-PR))Eq. (21)where PRG is the gerotor pump right side pressure, VRG is the right side of the gerotor pump control volume, ARG-R is rotary valve open area from port RG to port R, and PR is the right chamber pressure of the steering cylinders 456, 458.The right chamber pressure of the steering cylinders 456, 458 can be determined as:dPRdt=βVR⁢(Cd⁢ARG-R⁢2ρ⁢(PRG-PR)-x.s⁢As)Eq. (22)where Vg is the right side of the steering chamber control volume, xs is the steering piston displacement, and As is the piston area of the steering cylinders 456, 458 since they are symmetric.The left chamber pressure of the steering cylinders 456, 458 can be determined as:dPLdt=βVL⁢(x.⁢As-Cd⁢AL-T⁢2ρ⁢PL)Eq. (23)where PL is left chamber pressure of the steering cylinders 456, 458, VL is the left side of the steering chamber control volume, and AL-T is rotary valve open area from port L to port T.The rotary valve open areas for different ports are a function of the relative angle between spool and sleeve. The case that the spool stays at right side is similar to the description above.The torque Tge and the outlet flow Qge of the gerotor pump 452 can be determined as:Tge=Dge(PLG-PRG)2⁢π⁢ηgmEq. (24)Qge=Dge⁢θ.⁢ηgv2⁢π Eq. (25)Where Dge is the gerotor pump displacement, ηgm is the mechanical efficiency of the gerotor pump 452, and ηgv is the volumetric efficiency of the gerotor pump 452.The steering cylinders 456, 458 are connected to the steering valve as actuators. Two cylinders are cross connected. Change in the displacement of the cylinders 456, 458 will cause the corresponding change of wheel steering angles. The steering piston dynamics can be determined as:ms⁢x¨s=PR⁢AS-PL⁢AS-fs-bs⁢x.sEq. (26)where ms is mass of the steering pistons, fs is the frictional force of the steering pistons, and bs is the viscous coefficient of the steering pistons.The wheel steering angle is proportional to steering piston position from neutral position, and can be determined as:α=KxsEq. (27)where α is the wheel steering angle, and K is the gain from the steering piston position to the wheel steering angle.Vehicle ModelReturning to FIG. 5, the vehicle model combines the outputs of the drivetrain and the steering circuit to calculate the vehicle position and orientation. The vehicle's orientation in a 2-D plane is defined by the heading angle ψ, which can be determined as:ψ=2L⁢v⁢ tan⁢ α2Eq. (28)where L is the wheelbase length. The vehicle 2-D position dynamics can then be determined as:X.=v⁢ cos⁢ψEq. (29)Y.=v⁢ sin⁢ψEq. (30)Model ValidationThe models above can be validated using the control architecture automated wheel loader control system 300 of FIG. 3 (that otherwise reflect the control architecture 200 of FIG. 2). In this regard, and as described above, a key variable to be determined is the engine speed. The strategy of FIG. 4 can be applied. On the upper branch of FIG. 4, the required driveshaft speed ωf and drive torque Td can be determined by the reference vehicle speed and resistance forces through the driveshaft dynamics of Equations (5)-(8). Once the desired vehicle speed v is determined, the driveshaft speed ωf can be determined by:ωf=vkf / RwEq. (31)The final drive load torque Tf can be computed with Equation (7) since all parameters are known. With Equation (5), the drive torque Td can be determined by:Td=Jf⁢ω.f+Tf+TbEq. (32)With this approach, the brake torque Tb is assumed to be 0. Based on the shift schedule, the gear ratio γ can be determined. The turbine torque Ttc<sub2>t < / sub2>can be determined by:Ttct=Td / γEq. (33)The turbine speed can be determined by:ωtct=γωfEq. (34)The torque converter pump speed ωtc<sub2>p < / sub2>which is the same as the engine speed we, can then be determined using the torque converter model of Equations (2)-(4).EXAMPLESEmbodiments and advantages of features of the present disclosure are further illustrated in the following non-limiting examples. The operating conditions and details recited in these examples should not be construed to unduly limit the scope of the present disclosure.System validation was performed using field test data at the subsystem level and at the overall system level. For engine and drivetrain models, engine shaft dynamics and drive shaft dynamics were validated. For the hydraulic working circuit and the hydraulic steering circuit, the flows produced by the model were compared with experimental data with high accuracy. For this fully integrated system model, the difference between the measured fuel consumption of the loading cycle and the simulation result is 2.3%.The fully integrated and validated model could then be simulated for a typical loading cycle. A typical loading cycle includes six phases:Phase 1—Digging. The vehicle is at the material pile. The lift piston extends to lift the bucket and the tilt piston extends to load the material while the vehicle drives forward slowly.Phase 2—Transport Reverse 1. The vehicle drive backwards away from the pile and turns towards the truck.Phase 3—Transport Forward 1. The wheel loader drives forward to the truck while the lift piston extends to raise the full bucket.Phase 4—Dumping. The tilt piston retracts to dump the material into the truck.

[0089] Phase 5—Transport Reverse 2. The wheel loader drives backward to the starting point while the lift piston retracts to lower the empty bucket.

[0090] Phase 6—Transport Forward 2. The vehicle drives forward towards the pile of material.

[0091] The load forces during the loading cycle for lift and tilt cylinders are necessary for simulation. They can be determined as follows. Lift and tilt cylinder motion can be transformed to bucket motion and orientation through the linkage mechanism, and vice-versa. The interaction force between the bucket and the soil can be determined by the fundamental earthmoving equation during digging as described in Lee et al., “Force-balancing algorithm to remove the discontinuity in soil force during wheel loader excavation.”Journal of Mechanical Science and Technology, 32(10), pp. 4951-4957 (2018), the entire teachings of which are incorporated herein by reference. During the transportation before dumping, the constant weight of the material and bucket is the dominant force. The material weight decreases parabolically during dumping. The constant weight of the empty bucket is the dominant force for the rest of the transportation.

[0092] These bucket forces can then be converted to lift and tilt cylinder forces through the linkage mechanism. The force on the bucket acting against vehicle motion during digging is also the resistance force. The loading cycle external forces are shown in FIG. 9.

[0093] Simulation was conducted via MATLAB / Simulink. The tracking performance of the system compared to the references during Phases 1-6 is shown in FIG. 10 (as shown, the “Desired” plot line and the “Simulated” plot line are nearly identical). The Root-Mean-Squared-Error (RMSE) for the lift, tilt, vehicle, and engine speed tracking were found to be 3.1%, 0.8%, 4.9% and 4.3%, which indicates the control architecture is effective.

[0094] The tracking performance of the vehicle trajectory is shown in FIG. 11. The vehicle starts at point 1. From point 1 to point 2, the vehicle fills the bucket at the Pile. After that, the vehicle reverses to point 3 and moves forward to the Truck. At point 4, the vehicle dumps the material. The plot lines of FIG. 11 reflect that there is tracking overshoot during steep steering, but it settles down eventually. After dumping the material, the vehicle reverses from point 4 to point 5 and then advances to the Pile for the next cycle.

[0095] The torques acting on the engine shaft are shown in FIG. 12. The main pump torque is large during digging and bucket lifting, when the working circuit is doing work. The torque converter torque is never small and dominates the engine load during the transportation phases to accelerate the vehicle. It is also the highest during digging, since the resistance force induces a large load on the vehicle as it moves into the pile.

[0096] The pressure and flow for the hydraulic circuit is shown in FIG. 13. The main pump pressure is always larger than the highest working or steering circuit pressures unless the working circuit is load holding or lowering. During these cases, the pressure can be held constant or lowering without power from the pump.

[0097] The flow required by the working circuit is mainly during bucket filling and lifting, when there are cylinder motions. There is minimal flow to the working circuit during bucket lowering, when it uses regeneration. The flow of the steering circuit is relatively lower than the flow of the working circuit. The main pump provides the total flow to both circuits.

[0098] Energy flow analysis was performed based on the simulation results. The subsystem energy flow is shown in FIG. 14.

[0099] The first part of the energy flow from the engine is to the drivetrain. Specifically, the flow of energy in the drivetrain is from the engine to the torque converter, then through the transmission, and finally through the final drive shaft to the wheels. The torque converter takes 65.7% of the energy produced by the engine. 50.8% of the energy is transferred to the wheels to produce vehicle motion. The detailed drivetrain power analysis results are shown in FIG. 15. The torque converter gets most engine power during vehicle acceleration. The torque converter output follows a similar trend to the drivetrain input power. The torque converter loss is highest during digging. The reason is as follows. The vehicle speed is low at that time while the working circuit requires high pump flow, which means the engine speed is high. The ratio of the torque converter pump speed and turbine speed is high, so the torque converter operates at low efficiency.

[0100] Returning to FIG. 14, the second portion of the energy flow from the engine goes to the hydraulic circuit. Specifically, 22.7% of the energy from the engine flows to the main pump. The priority valve distributes the main pump output energy between the loader valve and the steering valve, with some throttling losses. Finally, energy flows into the lift and tilt cylinders to generate work tool motion, while the steering cylinders receive energy from the steering valve to turn the wheels. The detailed hydraulic circuit power analysis results are shown in FIG. 16. The energy loss of a subsystem is the difference between the input energy and the output energy. Main pump losses are mainly caused by low displacement operation or idling. The lifting loss is mainly the valve throttling loss during bucket lifting. The tilt loss during bucket filling is greater than the lift loss because the main pump provides high pressure due to the greater load force on the lift cylinder. The pressure drop from the main pump pressure to the tilt pressure is higher, which results in more throttling losses for the tilt valve. Steering losses are approximately half of the input energy due to throttling losses in valves and gerotor pump.

[0101] FIG. 17 shows the energy distribution across the entire cycle for the subsystems. Drivetrain transmits most energy from the engine. The energy consumption of the brake pump is approximately the same across the cycle. However, the relative energy consumption of the drivetrain and main pump is variable throughout the cycle. During transportation, especially during acceleration, the drivetrain accounts for a large portion of the engine output energy. While the main pump increases its share of energy consumption during bucket lifting. In addition, during digging, the main pump accounts for a larger portion of the engine output energy than any other time. It's worth noting that the drivetrain still consumes almost half of the engine's output energy during digging.

[0102] The systems and methods of the present disclosure provide a marked improvement over previous designs. The control system architectures of the present disclosure ensure both the drive function and the work function of an automated wheel loader can be operated in a desired manner while the engine can be operated to provide the required power for both functions in real-time. For example, the control architectures of the present disclosure decouple the MIMO system into several SISO systems, leading to the facilitation of designing controllers and implementing them. In other aspects, a fully integrated wheel loader model is provided, which includes engine, drivetrain, working circuit, steering circuit, and vehicle. Based on the model and the control architecture, a tracking problem is formulated with a typical loading cycle. The 2.3% difference between the measured and simulated fuel consumption verifies the accuracy of the model and the effectiveness of the control architecture. The control architectures and models of the present disclosure can facilitate the model-based control design and performance optimization for an autonomous wheel loader.

[0103] Although the present disclosure has been described with reference to preferred embodiments, workers skilled in the art will recognize that changes can be made in form and detail without departing from the spirit and scope of the present disclosure. For example, the control system architectures are useful not only for automated wheel loaders, but also other off-road vehicles.

Examples

examples

Embodiments and advantages of features of the present disclosure are further illustrated in the following non-limiting examples. The operating conditions and details recited in these examples should not be construed to unduly limit the scope of the present disclosure.

System validation was performed using field test data at the subsystem level and at the overall system level. For engine and drivetrain models, engine shaft dynamics and drive shaft dynamics were validated. For the hydraulic working circuit and the hydraulic steering circuit, the flows produced by the model were compared with experimental data with high accuracy. For this fully integrated system model, the difference between the measured fuel consumption of the loading cycle and the simulation result is 2.3%.

The fully integrated and validated model could then be simulated for a typical loading cycle. A typical loading cycle includes six phases:Phase 1—Digging. The vehicle is at the material pile. The lift piston extends...

Claims

1. An autonomous wheel loader control system comprising:a vehicle speed subsystem including a vehicle speed controller configured to control a speed of the wheel loader relative to ground;a first bucket subsystem including a first bucket controller configured to control at least one of a lift and a tilt of a bucket of the wheel loader; andan engine speed subsystem including an engine speed controller configured to control a throttle of the wheel loader;wherein the vehicle speed controller, the first bucket controller and the engine speed controller are decoupled from one another.

2. The system of claim 1, wherein the engine speed controller determines a desired setting for the throttle based, at least in part, upon an engine speed reference profile.

3. The system of claim 2, wherein the engine speed reference profile is derived from expected operating conditions of the vehicle speed subsystem and of the first bucket subsystem.

4. The system of claim 3, wherein the engine speed controller determines the desired setting for the throttle further based, at least in part, upon a current speed of the engine.

5. The system of claim 1, wherein the engine speed controller is programmed to determine a desired engine speed based upon a reference profile including a desired vehicle speed, a desired vehicle trajectory, a desired lift speed, and a desired tilt speed.

6. The system of claim 5, wherein the engine speed controller is further programed to determine an engine speed required by a drivetrain of the wheel loader and an engine speed required by a bucket hydraulic system of the wheel loader, and further wherein the desired engine speed is designated as a maximum value of the engine speed required by the drivetrain and the engine speed required by the bucket hydraulic system at each time instance in an operational cycle of the wheel loader.

7. The system of claim 1, wherein the first bucket controller is configured to control a lift of the bucket.

8. The system of claim 7, further comprising:a second bucket subsystem including a second bucket controller configured to control a tilt of the bucket.

9. The system of claim 8, wherein the second bucket controller is decoupled from the engine speed controller.

10. The system of claim 1, further comprising:a steering subsystem including a steering controller configured to control a steering angle of the wheel loader.

11. The system of claim 10, wherein the steering controller is decoupled from the engine speed controller.

12. The system of claim 1, wherein a reference profile for the vehicle speed controller is generated by an optimization algorithm.

13. The system of claim 12, wherein a reference profile for the first bucket controller is generated by an optimization algorithm.

14. The system of claim 13, wherein the first bucket controller is configured to control a lift of the bucket, the system further comprising:a second bucket subsystem including a second bucket controller configured to control a tilt of the bucket;wherein a reference profile for the second bucket controller is generated by an optimization algorithm.

15. The system of claim 14, further comprising:a steering subsystem including a steering controller configured to control a steering angle of the wheel loader;wherein a reference profile for the steering controller is generated by an optimization algorithm.

16. The system of claim 1, wherein the system is configured to operate the engine to provide required power for operation of the vehicle speed subsystem and the first bucket subsystem.

Images