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WO1996005975A1 - Systeme de suspension adaptative optimise par ordinateur et ameliorations apportees au procede associe - Google Patents

Systeme de suspension adaptative optimise par ordinateur et ameliorations apportees au procede associe Download PDF

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Publication number
WO1996005975A1
WO1996005975A1 PCT/US1995/010641 US9510641W WO9605975A1 WO 1996005975 A1 WO1996005975 A1 WO 1996005975A1 US 9510641 W US9510641 W US 9510641W WO 9605975 A1 WO9605975 A1 WO 9605975A1
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WO
WIPO (PCT)
Prior art keywords
compression
rebound
controller
motion
signals
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Ceased
Application number
PCT/US1995/010641
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English (en)
Inventor
James M. Hamilton
Lonnie K. Woods
Thomas P. Trotta
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
AimRite Systems International Inc
Original Assignee
AimRite Systems International Inc
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by AimRite Systems International Inc filed Critical AimRite Systems International Inc
Priority to AU34107/95A priority Critical patent/AU3410795A/en
Publication of WO1996005975A1 publication Critical patent/WO1996005975A1/fr
Anticipated expiration legal-status Critical
Ceased legal-status Critical Current

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Classifications

    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G17/00—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load
    • B60G17/015—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load the regulating means comprising electric or electronic elements
    • B60G17/0195—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load the regulating means comprising electric or electronic elements characterised by the regulation being combined with other vehicle control systems
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G17/00—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load
    • B60G17/015—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load the regulating means comprising electric or electronic elements
    • B60G17/018—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load the regulating means comprising electric or electronic elements characterised by the use of a specific signal treatment or control method
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G17/00—Resilient suspensions having means for adjusting the spring or vibration-damper characteristics, for regulating the distance between a supporting surface and a sprung part of vehicle or for locking suspension during use to meet varying vehicular or surface conditions, e.g. due to speed or load
    • B60G17/06—Characteristics of dampers, e.g. mechanical dampers
    • B60G17/08—Characteristics of fluid dampers
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2204/00—Indexing codes related to suspensions per se or to auxiliary parts
    • B60G2204/80—Interactive suspensions; arrangement affecting more than one suspension unit
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2400/00—Indexing codes relating to detected, measured or calculated conditions or factors
    • B60G2400/10—Acceleration; Deceleration
    • B60G2400/102—Acceleration; Deceleration vertical
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2400/00—Indexing codes relating to detected, measured or calculated conditions or factors
    • B60G2400/10—Acceleration; Deceleration
    • B60G2400/106—Acceleration; Deceleration longitudinal with regard to vehicle, e.g. braking
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2400/00—Indexing codes relating to detected, measured or calculated conditions or factors
    • B60G2400/20—Speed
    • B60G2400/204—Vehicle speed
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2500/00—Indexing codes relating to the regulated action or device
    • B60G2500/10—Damping action or damper
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2600/00—Indexing codes relating to particular elements, systems or processes used on suspension systems or suspension control systems
    • B60G2600/12—Sampling or average detecting; Addition or substraction
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2600/00—Indexing codes relating to particular elements, systems or processes used on suspension systems or suspension control systems
    • B60G2600/14—Differentiating means, i.e. differential control
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2600/00—Indexing codes relating to particular elements, systems or processes used on suspension systems or suspension control systems
    • B60G2600/16—Integrating means, i.e. integral control
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2600/00—Indexing codes relating to particular elements, systems or processes used on suspension systems or suspension control systems
    • B60G2600/76—Digital systems
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2600/00—Indexing codes relating to particular elements, systems or processes used on suspension systems or suspension control systems
    • B60G2600/90—Indexing codes relating to particular elements, systems or processes used on suspension systems or suspension control systems other signal treatment means
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2800/00—Indexing codes relating to the type of movement or to the condition of the vehicle and to the end result to be achieved by the control action
    • B60G2800/01—Attitude or posture control
    • B60G2800/012—Rolling condition
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2800/00—Indexing codes relating to the type of movement or to the condition of the vehicle and to the end result to be achieved by the control action
    • B60G2800/01—Attitude or posture control
    • B60G2800/014—Pitch; Nose dive
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B60—VEHICLES IN GENERAL
    • B60G—VEHICLE SUSPENSION ARRANGEMENTS
    • B60G2800/00—Indexing codes relating to the type of movement or to the condition of the vehicle and to the end result to be achieved by the control action
    • B60G2800/22—Braking, stopping

Definitions

  • the present invention relates to computer controlled vehicle suspension systems and methods, and more particularly to vehicle suspension systems and methods in which computer controlled damping forces in compression and rebound directions are used to optimize ride and handling characteristics of the vehicle.
  • Vehicle suspension systems have included shock absorbers, springs (coil, leaf, air or torsion bar), axle housings, torque arms, A-frames, anti-roll bars and stabilizers, and other elements. These components have been assembled in various combinations to produce the desired ride and handling characteristics for a particular vehicle. The characteristics of these devices often lead suspension system designers to compromise between various ride and handling characteristics and the design of such suspension systems. In a typical suspension system, changes in the spacing between the wheels/axles (unsprung mass) and the body /chassis (sprung mass) are cushioned by spring devices. Excess movement created by these spring devices must typically be controlled by damping devices.
  • the damping device is usually a velocity sensitive hydraulic system which uses hydraulic pressure to resist movement of a piston. Piston velocity is a direct function of the speed of suspension movement. The greater the piston velocity, the greater the force created in a direction opposite the movement. This force is created when the moving piston causes the hydraulic fluid, typically oil, to be forced through orifices and valves. The flow resistance encountered by the oil results in damping forces, both in compression and in rebound, and this damping acts to counter and dissipate the movement induced by the springs. Varying the fluid flow through the valves and orifices varies the forces acting against the spring-induced forces and, therefore, changes the ride and handling characteristics of the vehicle.
  • the damping forces are passive or resistive forces such that the respective compression and rebound forces only have effect when the damper is being contracted and extended, respectively.
  • the typical spring force is proportional to spring deflection.
  • the typical damping force is proportional to the velocity of the piston, that is, the velocity at which spring position changes.
  • Modern hydraulic suspension systems include numerous complexities to account for this difference between the sources of these two forces, position and rate of change of position. For example, a 6-stage valve system, 3 in compression and 3 in rebound, is known.
  • Vehicle leveling systems for maintaining predetermined height between the sprung mass of the vehicle (frame and body) and the unsprung mass (wheels, drive train, front axle and rear axle) are known. Many of these systems utilize air springs wherein air is pumped into or out of such air springs to modify the position of the vehicle body relative to the wheels.
  • air springs wherein air is pumped into or out of such air springs to modify the position of the vehicle body relative to the wheels.
  • Such systems are disclosed in U.S. Pat. Nos. 3,574,352, 3,584,893, 3,366,286, 3,830,138, 3,873,123, 4,017,099, 4,054,295, 4,076,275, 4,084,830, 4,162,083, 4,164,664, 4,105,216, 4,168,840, and 4,185,845.
  • These systems are not meant to actively adjust suspension system performance during vehicle travel to improve ride and handling characteristics.
  • Dynamic vehicle suspension systems have been developed. Systems for controlling the roll of a vehicle during a turn are disclosed in U.S. Pat. Nos. 2,967,062, 2,993,705, and 3,608,925. U.S. Pat Nos. 3,807,678 and 4,634,142 disclose active suspension systems which operate by changing the damping characteristics of the damping devices, based on a variety of discrete system states. These systems disclose utilizing a control system to respond to motion detected at the natural harmonic frequency of the suspension system to improve ride and handling characteristics. These control systems are not, however, optimum. These control systems may result in harsh handling characteristics or inadequate response to encountered road and driving conditions.
  • the damping devices damp movement between sprung and unsprung masses in both compression and rebound directions.
  • the control system utilizes a sensor at each damping device for generating position signals indicating the relative position (displacement) of the sprung and unsprung masses as related by the extension of the damping device. Based upon these position signals, the control system iteratively determines the desired damping forces to ensure optimum ride and handling characteristics.
  • the desired compression and rebound resisting damping forces are determined, in part, by determining the amplitude of motion at the natural frequency of the sprung mass system by using an appropriate filter and then by applying proportional damping forces in the direction opposite to the displacement of the spring from its normal (equilibrium) position.
  • a constant force is approximately maintained on the vehicle (emulating a constant force spring), thus minimizing body motions.
  • This force is calculated by multiplying the amplitude of motion at the sprung natural frequency by the actual amplitude signal defining the wheel to body displacement from its desired position, after performing an integration of the actual amplitude signal to filter out unwanted high frequency motions. Since the damping device is a dissipative device, in distinction to a true constant force spring, any natural frequency motions are automatically damped and an optimum control of the sprung natural frequency motion is achieved in a very simple and robust manner.
  • the amplitude of motion at the natural frequency of the sprung mass is determined through the utilization of a Discrete Fourier Transform (DFT).
  • the DFT provides a selective response to a specific frequency such that a damping response can be provided for lower frequency inputs but higher frequency inputs can be ignored, thus weighing lower motion frequencies greater than higher motion frequencies.
  • the DFT actually forms a narrow bandwidth filter that attenuates frequencies above and below the selected frequency that its parameters are set for.
  • the DFT in the particular implementation detects the sprung and unsprung natural frequencies of the vehicle.
  • the amplitude of damping provided by the control system is also minimized.
  • the rate at which the value of the DFT amplitude of motion decreases is limited, then the rate at which the damping decreases is also slowed. This is useful so that as damped oscillations become very small, the damping does not cease too quickly (since the DFT output is also becoming very small).
  • control forces are provided only when required and, even then, only at minimum levels required to control the sprung natural frequency motion.
  • the amount of harshness encountered by vehicle passengers when the vehicle hits bumps at substantially random frequencies is drastically reduced.
  • the response of the control system is a function of frequency, amplitude and time, thereby improving suspension system return to steady state conditions.
  • initial integration of the actual amplitude signal prevents the unwanted application of damping forces in response to high frequency bumps when the movements caused by such bumps are superimposed on body movements at the sprung natural frequency. This provides an additional reduction in the harshness encountered by passengers.
  • an independent response in the compression and rebound directions such that the response is proportional to the extension of the damping device from its long-term average position provides improved ride and handling characteristics.
  • FIG. 1 is a diagram of a suspension system embodying the present invention
  • FIG. 2 is a diagram of a vehicle having a computer controlled suspension system and embodying the present invention
  • FIG. 3 is a block diagram of the computer control system and embodying the present invention.
  • FIG. 4 is a cross-sectional view of a damper unit suited for use in a suspension system and embodying the present invention
  • FIG. 5 is a simplified definitional diagram of a damper
  • FIG. 6 is a flow diagram of the computational process used by the computer control system of FIG. 2 and embodying the present invention
  • FIGS. 7a-f are block diagrams of preconditioning calculations of the process of FIG. 6;
  • FIGS. 8a-g are block diagrams of the sprung natural frequency response calculations of the process of FIG. 6;
  • FIGS. 9a-c are block diagrams of the unsprung natural frequency response calculations of the process of FIG. 6;
  • FIG. 10 is a block diagram of the pumping down control response calculations of the process of FIG. 6;
  • FIGS, l la-e are block diagrams of the roll control response calculations of the process of FIG. 6;
  • FIGS. 12a-e are block diagrams of the pitch control response calculations of the process of FIG. 6;
  • FIG. 13 is a block diagram of the topping and bottoming out control response calculations of the process of FIG. 6;
  • FIGS. 14a and b are block diagrams of the stored energy response calculations of the process of FIG. 6;
  • FIG. 15 is a block diagram of the float control calculations of the process of FIG. 6;
  • FIG. 16 is a block diagram of an unsprung natural frequency calculation of the process of FIG. 6;
  • FIG. 17 is a block diagram of a falling edge integrator calculation of the process of FIG. 6;
  • FIG. 18 is a block diagram of an alternate computer control system and embodying the present invention.
  • FIG. 19 is a block diagram of an alternate roll control response calculation of the process of FIG. 6.
  • FIG. 20 is a block diagram of an alternate pitch control response calculation of the process of FIG. 6.
  • a wheel 2 engages a road surface 1 and is rotatably mounted on an axle 4 which extends from one end of a carrier 6.
  • the other end of the carrier is pivotally mounted to a frame or chassis and body 7 of a vehicle 8.
  • a suspension unit 10 is connected between the chassis and axle.
  • the suspension unit combines a damping device 12 and a spring 14, which can be an air spring, coil spring, or similar device.
  • the wheel, axle, and carrier comprise the unsprung mass of the vehicle, and the frame and body comprise the sprung mass of the vehicle.
  • damping forces FQ and FR in compression and rebound, respectively, of the damping device are varied by a controller 16, which incorporates the novel control features of the present invention, to optimize the ride and handling characteristics of the vehicle under a wide range of driving conditions and eliminate harshness associated with prior systems.
  • the vehicle 8 comprising a steering mechanism 20, a hydraulic braking mechanism 22, four wheels, left and right front and left and right rear, 2a, 2b, 2c, 2d, respectively, suspended from the vehicle by suspension units as hereinabove described, and the controller 16 preferably a single controller associated with all four wheels.
  • An exemplary vehicle is a 6000 lb. truck for which a specific embodiment of a suspension system is described.
  • a damping device 12a-d, such as the damping device of FIG. 4, and a spring 14a-d couple each wheel, 2a-d respectively, to the body of the vehicle.
  • the controller 16 receives position data (P A/D ) via lines 17a-d from position sensors 190a-d on each one of the damping devices 12a-d, respectively, indicating me position of the extensible member of each damping device.
  • the subscript " A/D " indicates that the particular variable reflects raw, real-time data which is the output of an analog to digital converter associated with a sensor device (not shown) in each one of the sensors 190a-d.
  • Other sensing mechanisms that indicate wheel to body displacement that provide equivalent data could be used but may not be as cost effective.
  • the controller 16 also receives brake pressure ( ⁇ A/D ) via lnie 21, steering angle ( ⁇ A/D ) via line 19, and vehicle speed (S A/D ) via line 23 from sensors 28, 30, and 32, respectively.
  • the controller uses this data to determine the desired damping forces in the compression and rebound directions for application by the damping devices at each wheel. In this iterative manner, the controller uses such determinations to generate damper control signals for each damping device, to effect the desired damping forces in the compression and rebound directions, respectively.
  • the signals are delivered to the dampers 12a-d via lines 25a-d and 27a-d, respectively, for compression and rebound.
  • this continuous iterative process repeats at a fixed cycle or interval ⁇ (referred to hereinafter as “cycle” or “iteration”), which may be, by way of example, every 5 milliseconds, based upon a computer driven interrupt timer in the controller, although one skilled in the art would realize that a variety of timing mechanisms and frequencies are possible.
  • cycle or “iteration”
  • n- 1 indicates the iteration immediately preceding the present iteration denoted by "n”.
  • Exemplary controller 16 is a microcomputer having a microprocessor such as a Motorola model 68332 processor with 16 MHz/16 bit performance as well as appropriate ROM, RAM, power supply and related circuity.
  • An exemplary brake pressure sensor 28 is an EPNM-38-1000G by Entran.
  • the various sensors with analog outputs preferably have analog output signals which are converted to digital form for the controller to perform the required calculations.
  • a series of Analog-to-Digital (A/D) converters are provided for this purpose as described above.
  • Analog-to-digital (A/D) converters are associated with each sensor. Such converters can, optionally, be included in the computer which serves as the controller.
  • the output of the brake pressure sensor is monitored by a unipolar 8 bit minimum resolution A/D converter, providing a resolution of 4 PSI per least significant bit (LSB).
  • Exemplary steering angle sensor 30 can be in the form of a linear potentiometer attached to a control arm or other member of the vehicle steering gear.
  • a suitable potentiometer is the "M-Series" by Maurey Instruments Corp. of Chicago, Illinois.
  • the output of the potentiometer is monitored by a bipolar 8 bit minimum resolution A/D converter 29 ( ⁇ 128 levels), providing a resolution per LSB of 0.5 degrees of wheel turn.
  • an optical encoding system can be used wherein, for example, an optical sensor detects an array of lines or other markings on a steering column for detecting rotation of the column.
  • Exemplary speed sensor 32 is a Hall-effect sensor which detects a magnet connected to a drive shaft or other powertrain component. Many modern vehicles include some form of speed sensor for producing digital speed information, often associated with a vehicle transmission or with an anti-lock braking system (in which case a sensor is present for each wheel).
  • the A/D converter 33 is similar to converter 29 with a resolution of 1 mph/LSB.
  • the position sensors 190a-d can each comprise a linear position transducer connected to a hydraulic reservoir of the damper for measuring fluid displaced by the damper shaft as the damper is compressed.
  • a suitable transducer is the Series 240 (such as model 0243-0000) by Trans-Tek Inc. of Ellington, Connecticut.
  • the output of the potentiometer is monitored by bipolar 12 bit A/D converters 26a-d (shown in FIG. 2) ( ⁇ 2048 levels), thereby providing a resolution of 2/1000 inch per LSB for a sensor stroke range of ⁇ 1 inch.
  • the actual resolution used is 1/100 inch per LSB as referenced at the wheel, which allows for a larger range of input to cover expansion of the oil in the damper over wide automotive temperature ranges (typically -40° to +250°F). This expansion can be equivalent to 3-5 times the actual displacement of the damper shaft. This temperature range does not present a significant problem, however, since the resolution of the input P A/D allows sufficient range and the long term integration and adjustment used to calculate an adjusted position P compensates for the offset.
  • Each of dampers 12a-d has a corresponding signal converter 18a-d, respectively, for converting digital compression control and rebound control signals from computer or processor 15 to pulse width modulated (PWM) signals for control of regulators 40a and 40b in each damper and for converting analog position signals from the position sensors 190a-d in dampers 12a-d, respectively, to digital position signals for the computer or processor 15.
  • PWM pulse width modulated
  • Each of signal converters 18a-d is essentially identical and, therefore, only signal converter 18a will be described by way of example, it being understood that the others are alike.
  • Digital compression control and rebound control signals are converted from digital to analog and then to pulse modulated signals for compression and rebound regulators 40a and 40b; respectively, by digital to analog converters 52 and pulse width modulator circuits 24.
  • Analog position signals from position sensor 190a is converted to a digital signal by analog to digital converter 51.
  • the controller 16 receives data from sensors 32, 30, and 28, respectively, corresponding to the vehicle speed (S A/D ), brake pressure (P A/D ), and steering angle ( ⁇ A/D ) of the vehicle. Additionally, the controller receives from the position sensors 190a-d for each wheel 2a-d, respectively, through analog-to-digital converters 26a-d, the positions (P A/D ) 17a-d or extensions of the extensible member of the damping devices 12a-d at that wheel, such position or extension indicating the wheel to body displacement. The controller 16 utilizes these data inputs to determine the appropriate and desired damper forces in the compression and rebound directions for each wheel.
  • the controller 16 has input ports for receiving signals from sensors, output ports for providing control signals, and a central processor or computer 15.
  • the computer includes the standard elements such as CPU, RAM, ROM and the like for processing the input signals to provide the desired outputs.
  • the input and output ports can be formed as separate receiver and output portions of the controller or can be appropriately integrated within the computer.
  • the output signals from the controller to the pressure regulators in each damper 12a-d must be converted from digital form to a form appropriate for their correct operation.
  • Preferred regulators 40a-b as shown in FIG. 4 are controlled by the amount of electrical current received. Although a controlled current can be achieved by providing the regulators with an adjustable current or voltage source, this type of analog circuitry wastes a considerable amount of power as heat.
  • the output signals of the controller are received by pulse- width modulator (PWM) converter circuitry 24 (FIG. 3) which receive converted control signal from digital-to-analog converters and, responsive thereto provides a Pulse-Width-Modulated (PWM) signal that is then applied to each of regulators 40a-d for controlling that regulator for adjusting the force applied by the damper in compression or rebound, respectively.
  • PWM pulse- width modulator
  • the PWM converter circuity 24 uses little power by pulsing the voltage to the regulators ON and OFF with varying duty cycles to achieve varying currents. The power is saved because the PWM voltage switching mechanism requires a minimum of power. All of the above circuit techniques are well known in the art. It should also be appreciated that other forms of analog (or digital) devices could be provided for the regulators, depending upon the particular regulator implementation used.
  • damping device 12 A specific embodiment of the damping device 12 is shown in FIG. 4, the damping device being configured for use in an automotive or similar vehicle shock absorbing system.
  • the structure and operation of the damping device is discussed in further detail in our co- pending application Ser. No. 08/272,208, filed July 8, 1994 and international application number US 95/08550, filed
  • the damping device 12 comprises an extensible member or shaft 160 within a body or case formed of concentric inner and outer sleeves 150 and 152.
  • the extensible member axially reciprocates inside the case forcing hydraulic fluid, such as oil, through a first (compression) relief valve 40a and a second (rebound) relief valve 40b, which relief valves regulate pressure in the compression and rebound directions, respectively.
  • the blow-off pressures of these valves may be adjusted by the controller through the application of various voltages to solenoids 116a and 116b, thereby altering the compression and rebound damping forces applied by the damping device between the sprung and unsprung masses.
  • a port 180 is formed in the outer sleeve of the second valve 40b.
  • a hose 182 connects the port to a cylinder 184 in which rides a piston or extensible member 186.
  • a linear position transducer 190 is connected to the piston 186 for measuring its displacement caused by the displacement of fluid within the damper due to the varying length of the portion of the extensible member within the damper body.
  • the signal indicative of the position (P A/D ) of the extensible member is provided by the transducer 190 to the control system via wiring pairs 192 which can coincide with wiring 17a-d for the four wheels, only one wire being shown for each of 17a-d for simplicity.
  • An exemplary solenoid 116a and 116b is the model PS-16 proportional solenoid by Elwood of Milwaukee, Wisconsin.
  • An 8 bit digital-to-analog converter (not shown in FIG. 4) (256 levels) and PWM driver (shown in FIG. 3) is associated with the solenoid for providing necessary analog voltage based on the digital control signals from the controller.
  • Such a converter may, optionally, be included in the microcomputer which serves as the controller 16.
  • FIG. 5 shows the damping device 12 in its maximally compressed (retracted) state 60 (solid lines).
  • the maximally extended state 62 and long term average or equilibrium (normal) state 61 are shown in broken lines.
  • the received position signal P A/D is also pictorially shown, indicative of the displacement of the extensible member relative to a zero position 63 of the sensor.
  • the signal is positive while the suspension is compressed and negative while the suspension is extended.
  • a baseline position signal P LT1 is determined as a long term average, using a long term integration.
  • the adjusted position signal P (relative to equilibrium) is then P A/D - LT1. Similar adjustments are used for the outputs of the other sensors.
  • FIG. 6 is a schematic block diagram which shows the general steps of the computational process utilized by the computer controller to execute the control system of the present invention. The sequence is controlled by a computer program, preferably firmware in the computer controller. Calculations as discussed herein are performed by the computer under program control.
  • preconditioning (baseline) calculations (Block 522), more completely described hereinbelow in FIGS. 7a-f, are performed by the controller to provide the vehicle adjusted (baseline) speed S (FIG. 7e), steering angle ⁇ (FIG. 7a), brake fluid pressure p (FIG. IT), position P (FIGS. 7b-c) and velocity V p (FIG. 7d) of the extensible member of each damping device.
  • the various compression and rebound responses for each damper are calculated for a plurality of response modes and are more completely described hereinbelow in FIGS. 8-20.
  • the response modes include a Sprung Natural Frequency response F CSNF and F RSNF (Block 524), an Unsprung Natural Frequency response F CUNF and F RUNF (Block 526), a Pumping Down response F CPD and F RPD (Block 528), Pitch and Roll responses F CPH and F RPH an F CR anc F RR (Blocks 532 and 533), a Topping Out and Bottoming Out response (Block 534), a Stored Energy response F CSE and F RSE (Block 536), and a Float response F CFL and F RFL (Block 538). Finally, the responses are summed (Block 540) to yield the overall compression and rebound responses F C and F R for each damper.
  • the response modes are now defined.
  • the term "sprung natural frequency” identifies the frequency at which the sprung mass (the frame and body) tends to oscillate on the springs. This can be determined empirically for a particular vehicle, but for a typical car or light truck is on the order of a 1.5 Hz frequency.
  • the SNF response calculations (Block 524) determine the damping forces in the compression and rebound directions which will damp the vehicle motion due to spring oscillation. As more completely described hereinbelow in FIGS. 8a-g, this is done, in part, by determining the amplitude of body to wheel motion, as represented by the amplitude of motion of the extensible member at the sprung natural frequency, and by determining the current position of the extensible member. This information is used to provide a force proportional to and counteracting against the spring force inducing oscillation.
  • unsprung natural frequency identifies the frequency at which the unsprung mass (the wheels, axles and carriers) tends to oscillate on the springs between the sprung mass and the road surface. As with the SNF, this can be determined empirically, but for a typical car or light truck is on the order of a 10-15 Hz frequency.
  • the UNF response calculations (Block 526) determine the damping forces in the compression and rebound directions which will damp the wheel motion due to spring oscillation. As more completely described hereinbelow in FIGS. 9a-c, this is done, in part, by determining the amplitude of body to wheel motion, as represented by the position of the extensible member at the unsprung natural frequency, and providing a counterforce to damp this movement.
  • the term “pumping down” identifies the situation wherein the shock absorber compression forces during rapid wheel movements are less than the rebound forces, such that over an interval of time the net or total resulting force on the chassis or body is predominantly downward, pulling the chassis lower to the ground.
  • the rebound force exceeds the compression force of the damper, the rebound force hinders the return of the springs from a compressed state upon oscillation of the vehicle and thus decreasing the average extension of the damper during the oscillation.
  • the Pumping Down response calculations (Block 528) determine when the chassis is being pulled downward and decreases the rebound resisting force to counteract the pumping down effect caused by otherwise excessive rebound forces resisting return of the compressed spring.
  • bottoming out identifies the condition wherein a road perturbation, such as a bump or other influence, causes the wheel and axle to reach the upper limit of the dynamic range of travel of suspension, that is, the suspension system is compressed or retracted to its maximum limit.
  • the term “topping out” refers to the condition where a road perturbation, such as a hole or other influence, causes the wheel and axle to reach the lower limit of the dynamic range of travel of the suspension, that is, the suspension system is extended to its maximum limit.
  • the Bottoming and Topping Out (B/T) response calculations (Block 534) determine the damping forces required to avoid the occurrence of the aforementioned situations. As more completely described hereinbelow in FIG. 13, this is done, in part, by using the current position of the extensible member, the rate of change and the remaining travel from that position to determine the desired compression and rebound forces to avoid bottoming or topping out.
  • roll and pitch refer to rotational body movement, about respective longitudinal and transverse axes of the vehicle, due to respectively lateral (transverse) and longitudinal accelerations caused by turning (roll) and braking or hard accelerating (pitch). Braking tends to force down the front of the vehicle while raising the rear of the vehicle
  • Roll response calculations determine the damping forces required to counteract this body movement. As more completely described hereinbelow in FIGS, 11a-e, this is done, in part, by using sensors which indicate that at least one of roll or pitch is being induced and is otherwise imminent. Counteracting forces are then provided with the dampers to attempt to prevent corresponding roll or pitch motions, including feedback stabilization.
  • the positions of the extensible members about the transverse and longitudinal body axes are compared to determine the desired compression and rebound damping forces. For example, to determine the desired compression and rebound forces for the damper associated with the left front wheel, the left front wheel extensible member position is compared to the right front wheel extensible member position for roll response and to the left rear wheel extensible member position for pitch response.
  • An alternate pitch calculation comprises a comparison of diagonally opposite wheels.
  • the term "stored energy” identifies body movement along a diagonal axes of the body due to non-symmetrical road perturbations, that is, those which affect only a single wheel or combination of wheels at one end or side of the vehicle.
  • the Stored Energy response calculation (Block 536) determines the damping forces required to counteract this body movement. As more completely described hereinbelow, this is done, in part, by comparing the positions of the extensible members of diagonally placed (opposite) wheels to determine the desired compression and rebound damping. For example, to determine the desired compression and rebound forces for the left front wheel, the left front wheel extensible member position is compared to the right rear wheel extensible member position. If the left front wheel hits a large bump, the right rear damper applies a counteracting compression force to prevent the vehicle from rotating downward in the rear about its center of roll in response to the upward force exerting on the front (left front wheel) via the compressed left front spring.
  • float identifies the movement which the sprung mass (the frame and body) tends to experience in the form of small rotations in roll and pitch and vertical motion. These motions and rotations, although uncomfortable to passengers during extended driving, are often too small for previous algorithms to detect and control.
  • the specialized Float response calculations (Block 538) determine the damping forces which damp these subtle motions. As more completely described hereinbelow in FIG. 14, this is done, in part, by determining the amplitude of vertical, roll and pitch body motions, and using this information to provide a small "Coulomb” (constant) force that damps the oscillations. For each mode, the responses are calculated for both the compression and rebound directions.
  • the preconditioning calculations for the steering angle provide an adjusted and clipped steering angle.
  • the steering angle position sensor is calibrated to provide a zero reading for straight ahead motion
  • the data received from the sensor is adjusted to reduce error due to miscalibration or sensor drift.
  • the steering angle is also set to zero at low angles to additionally reduce error because the roll forces generated by small angle turns do not materially affect the ride and handling characteristics of the vehicle.
  • N ⁇ corresponds to a selected number of cycles equalling a set period of 0.64 seconds.
  • the magnitude of the received steering ⁇ A/D is compared (Diamond 710) to the magnitude of a constant angle ⁇ LIM' 5 degrees in the described embodiment. If ⁇ A/D has a magnitude greater than or equal to the magnitude of ⁇ L IM' ⁇ LTI remains unchanged (Block 712). The comparison of Block 710 is performed to avoid attempting to adjust for errors in the calibration of the sensor during high magnitude turns. If ⁇ A/D has a magnitude less than the magnitude of ⁇ LIM' ⁇ A/D iS compared to the previous long-term integrated steering angle ⁇ LTIn-1 (Diamond 714).
  • ⁇ A/D is greater than ⁇ LTin-1 ' ⁇ LTI is increased by a small amount ⁇ ⁇ (Block 716), 0.5 degrees in the described embodiment. If ⁇ A/D is less than ⁇ LTIn-1 ' ⁇ LTI is decreased by the small amount ⁇ ⁇ (Block 718). If ⁇ A/ D is equal to LTIn-1 ' ⁇ LTI is not changed (Block 707).
  • the adjusted steering angle ⁇ is then calculated by subtracting the long-term integrated steering angle ⁇ LTI from the received steering angle ⁇ A/D (Block 720). ⁇ is further set to zero (Block 724) if the magnitude of ⁇ is less than a minimum value ⁇ CLIP (Diamond 722), by way of example, 1 degree in the described embodiment.
  • FIG. 7b shows the preconditioning calculations for the long term integrated equilibrium position P LTI of each of the extensible members.
  • the position P A/D is received (Block 726).
  • the previously calculated ⁇ (FIG. 7a) is compared to zero (Diamond 727) to prevent the long term integrated position P LTI from changing when it should not. For example, cornering results in temporary changes in position P A/D that do not represent the average normal ride position desired.
  • a counter (Diamond 728 and Blocks 729 and 730) is used similar to that used in the preconditioning calculations for 0 in FIG.
  • the received extensible member position P A/D is compared (Diamond 734) with the previous long-term average position value P LTIn-1 . If the received extensible member position P A/D is greater than P LTIn-1 , P LTI is increased by adding a small amount ⁇ p , such as 1/100 inch in the described embodiment, to the previous value (Block 736).
  • the adjusted position relative to equilibrium of the extensible member P is calculated.
  • P A/D is received (Block 739) and the long term integrated equilibrium position P LTI is subtracted from P A/D (Block 740).
  • N P can be 1, if the integrated rate must be slow, the addition or subtraction of the smallest amount during each cycle might still be too fast.
  • a counter is implemented so that a small amount is added after a repeated interval of a larger number of cycles Np. For example, an integration of P LTI in the described embodiment adds or subtracts 1/100 inch every 50 cycles, each cycle being 5ms, thus having a 250ms delay between increments.
  • FIG. 7d shows, for each extensible member 17a-d, the calculation of damper velocity V P .
  • the current received position P A/D (Block 755) is received.
  • the previously received extensible member position P A/Dn-1 is subtracted from the current received position P A/D and the difference is divided by the cycle ten (Block 756).
  • the rate of change V P of the position of the extensible member is calculated as an aid to the hereinbelow described Bottoming and Topping Out response (Block 534).
  • FIG. 7e shows the determination of the integrated speed S of the vehicle which provides an improved controller response.
  • the speed S A/D is received (Block 757).
  • the received vehicle speed S A/D is compared to the integrated speed S n-1 (Diamond 761). If the received vehicle speed S A/D is greater than the previous integrated speed S n-1 , then the integrated speed S is increased by a small amount ⁇ S (Block 762). If the received speed S A/D is less than the previous integrated speed S n-1 , then the integrated speed S n-1 is decreased by a small amount ⁇ S (Block 764).
  • S A/D is equal to S n-1 , then S is not changed (Block 766).
  • the small amount ⁇ S is 1 MPH every 500ms to achieve the correct integrated rate.
  • the preconditioning calculations for the brake pressure ⁇ are similar to the preconditioning calculations for the steering angle ⁇ , however, brake pressure is assumed to be positive. In embodiments wherein an indication of acceleration that can be used to predict both dive and squat is present, such as with an accelerometer, the parameter for brake pressure would have both positive and negative values. As shown in FIG. 7f, the pressure
  • P A/D is received (Block 770) at a set period, 250ms in the described embodiment as determined by a counter (Diamond 772 and Blocks 774 and 776), which corresponds to a selected number of cycles N ⁇ of the system equal to 50.
  • the magnitude of the received brake pressure ⁇ A/D is compared (Diamond 778) with the magnitude of a constant pressure ⁇ LIM' 25 psi in the described embodiment. If ⁇ A/D has a magnitude greater than or equal to the magnitude of the constant pressure ⁇ LIM' then the long-term integrated brake pressure ⁇ LTI will not be changed (Block 780).
  • ⁇ A/D has a magnitude less than the magnitude of the constant pressure P LIM' the n ⁇ A/D is compared with the previous long-term integrated brake pressure ⁇ LTI n-1 (Diamond 782). If ⁇ A/D is greater than ⁇ LTIn- 1 ' ⁇ LTI is increased by a small amount ⁇ ⁇ (Block 784), 1 psi in the described embodiment. If ⁇ A/D is less than ⁇ LTIn-1 ' ⁇ LTI is decreased by the small amount ⁇ ⁇ (Block 786). If ⁇ A/D is equal to ⁇ LTIn- 1 then ⁇ LTI is not changed (Block 775).
  • the adjusted brake pressure ⁇ is then calculated by subtracting the long-term integrated brake pressure ⁇ LTI from the received brake pressure ⁇ A /D (Block 790). ⁇ is further set to zero (Block 794) if the magnitude of ⁇ is less than a minimum value ⁇ CLIP (Diamond 792), by way of example, 10 psi in the described embodiment.
  • ⁇ CLIP Minimum Value
  • FIG. 8a shows the steps of calculating the Sprung Natural Frequency response (Block 524) which are subsequently described in detail in FIGS. 8b-g.
  • An integration or filtering of the extensible member position relative to equilibrium, P is performed (Block 802), resulting in an integrated position P SNF of the extensible member.
  • Such an integration can be either a linear or non-linear integration, with the described embodiment, by way of example, using a non-linear integrator to increase controller response to large perturbations.
  • This integration allows movement at the sprung natural frequency to be detected, while filtering out unwanted movements at higher frequencies. Any common low pass or bandpass filter technique can be used to detect the sprung natural frequency.
  • a Discrete Fourier Transform is performed on the integrated position P SNF (Block 804), resulting in an intermediate parameter A DFTSNF , and thereby providing the amplitude of the motion of the extensible member at a selected frequency, that is, the sprung natural frequency.
  • the intermediate parameter A DFTSNF is modified by being multiplied by a vehicle dependent gain factor and a vehicle speed dependent gain factor to increase the controller response at higher speeds, which is also clipped or limited to a maximum value (Block 806), yielding A SNF .
  • a SNF is smoothed (Block 808) to yield an integrated parameter A INTSNF , an operation which is later described in detail, when the magnitude of the intermediate parameter A DFTSNF is decreasing.
  • the smoothing results in increased controller response as motion is damped.
  • the integrated parameter A INTSNF is then multiplied by both the absolute value of P SNF and a vehicle dependent gain factor (Block 810), resulting in an intermediate parameter proportional to the amplitude of motion of the extensible member at a selected frequency and the current displacement of the extensible member.
  • This parameter is then used to provide forces which are proportional to and in an opposite direction to the spring force F RSNF and F CSNF , respectively.
  • the Roll and Pitch algorithms are already providing forces, only the amount exceeding such already- existing forces is provided.
  • FIG. 8b shows the details of the non-linear integration (Block 802) of the position of the extensible member. Every iteration, the position of the extensible member P is compared to the value of the previously calculated integrated position value P SNFn-1 (Diamond 820). If P is greater than P SNFn-1 then an integrated value is added to P SNFn-1 (Block 822). In the described embodiment, the integrated value being, by way of example, K SNF
  • P SNF P SNFn-1 + K SNF
  • FIG. 8c shows that a Discrete Fourier Transform (DFT) is performed (Block 804) on the integrated position value P SNF resulting from the previous step of FIG. 8b.
  • DFT Discrete Fourier Transform
  • FIG. 8d shows the vehicle speed dependent gain factor K SSNF being calculated as a function of the form S ⁇ S REFSNF , where S REFSNF is a constant reference speed for which the speed dependent gain factor is 1 and S is the integrated speed of the vehicle.
  • the value of this gain factor is not allowed to go below a minimum value, such value being determined by the value at an empirically determined specific vehicle speed S MIN .
  • S > S MIN (Diamond 844)
  • K SSNF is set to S ⁇ S REFSNF (Block 846)
  • K SSNF is set to S MIN ⁇ S REFSNF (Block 848).
  • the values are as follows:
  • the result A DFTSNF of performing the DFT is multiplied (Block 834) by the vehicle speed dependent gain factor K SSNF (calculated above in FIG. 8d) and a constant vehicle dependent gain factor GAIN SNF , which are both based on the characteristics of the vehicle.
  • the result (Block 806) of this multiplication A SNF is limited to a maximum value A DFTSNFMAX , 1.0 inch in the described embodiment (Block 838), if the result of the multiplication would otherwise result in a greater value (Block 836).
  • FIG. 8f shows that A SNF is smoothed when its value decreases, resulting in A INTSNF (Block 808).
  • This smoothing process helps maintain the damping level up as the motions are decreased so as to quickly and effectively stop the body motions. This is done every N A cycles (1 in the described embodiment) determined by counter C A (Diamond 850 and Blocks 852 and 854).
  • the value of A SNF is compared to the previous value of A INTSNFn-1 (Diamond 856) and A INTSNF is set to the value of A SNF if A SNF is greater than or equal to the previous value of A INTSNF (Block 851).
  • a SNF is less than the previous value of A INTSNF , however, then a constant value ⁇ A , 1 in the described embodiment, is subtracted from the previous value of A INTSNF to form a new value of A INTSNF (Block 858).
  • is multiplied by A INTSNF and either a rebound constant (Block 860) or a compression constant (Block 862) to form the compression and rebound desired forces F CSNF and F RSNF (Block 810).
  • a rebound constant Block 860
  • a compression constant Block 862
  • Exemplary values for these constants are varied for the front and rear wheels of the vehicle and are listed below (wherein K CSNFF indicates the front compression constant, K RSNFR indicates the rear rebound constant, and so forth):
  • the constants above are dependent on the resolution of the inputs and outputs of the overall system.
  • An output value of 1 results in an approximate damping force of 7 pounds. Therefore, as an example, if the above mentioned constant equals 1 and the vehicle is moving at a sprung natural frequency with an amplitude of 1 inch, the algorithm provides an output of 100 (1.0 constant X 100 DFT amplitude) which results in a damping force of approximately 700 pounds (100 X 7 pounds).
  • U.S. patent 4,634,142 discloses that a filter can be used to detect when the sprung mass is starting to oscillate. The filter approach, however, was dismissed because the output of a filter has a delay and it was felt that the response would be too slow and the performance would be inadequate. Furthermore, the action that would be required to damp the mass if the filter was used is not described (that is, if there is an output indication of natural frequency motion, then how would that be utilized to damp the mass). It was found that the algorithm (and hardware) disclosed in the '142 patent controlled the mass very well. However, the ride was too harsh over bumps faster than the natural frequency of the masses because the displacement using signals only proportional to displacement of the masses damped all of the time, even for fast bumps above the natural frequency.
  • modulating the displacement-only approach in the '141 patent with a filter as in FIG. 8c and 8g provides a ride performance that is improved significantly. Additional improvements include: 1) modulation of the forces by vehicle speed as in FIG. 8d; 2) pre-filtering of position to eliminate high frequency inputs to the DFT as in FIG. 8b; 3) limiting the maximum control forces as in FIG. 8e; and 4) a falling edge integrator to hold the damping force up as the oscillations damp out (and the output of the DFT falls) as in FIG. 8f.
  • the controller determines the response to wheel movement at the unsprung natural frequency in a manner similar to that of the sprung mass response.
  • a filter can detect the unsprung natural frequency.
  • a DFT is performed (Block 902) on the adjusted extensible member position P to determine the amplitude of motion P DFTUNF at the unsprung natural frequency.
  • the adjusted extensible member position data is not filtered, as was done for the sprung natural frequency response, due to the negative effect it would have on determining the amplitude of the higher frequency unsprung motion.
  • a typical sprung natural frequency is on the order of 1 Hz
  • the unsprung natural frequency is on the order of 10 Hz.
  • the amplitude of motion P DFTUNF can be optionally smoothed, in the same manner as was described for A INTSNF above, when the current value of
  • P DFTUNF has a smaller magnitude than the previous value. This is not required and not included in the described embodiment.
  • Low amplitude motion at the unsprung natural frequency is optionally clipped and filtered (Block 908), however, such motion does not have appreciable effects on ride and handling characteristics but does reduce harshness on rough roads.
  • the output of the DFT is then multiplied by a vehicle dependent gain factor (Block 910) to yield a desired rebound force.
  • the DFT shown in FIG. 9b, is similar to that for the SNF mode but is performed on the previous 15 iterations. 15 cycles at 5ms per cycle represents 75ms which represents approximately 13 Hz -- the desired frequency to detect in the described embodiment.
  • FIG. 16 An alternative and preferred approach to controlling the unsprung natural frequency is shown in FIG. 16.
  • a filter such as a DFT (Block 1610) determines if the unsprung mass is moving at its natural frequency, A DFTUNF .
  • This signal is then smoothed with a falling edge integrator (Block 1611) to hold the damping force up as the oscillations damp out.
  • the integrator is shown in FIG. 17 and is essentially the same as the one for SNF as shown in FIG. 8f.
  • the desired rebound damping force is then determined by multiplying the integrated DFT signal A INTUNF , determined at Blocks 1675 and 1685 of FIG. 17, by the damper velocity V p determined in FIG.
  • N UNF in FIG. 17 is 1 and K UNF is 1.
  • FIG. 9c The details of the clipping and filtering of low amplitude motion is shown in FIG. 9c.
  • the amplitude of motion A DFTUNF received from the DFT is compared (Diamond 920) with a first constant value, 10/100 inch by way of example in the described UNF CLIP embodiment. If the amplitude of motion A DFTUNF is greater than UNF CLIP , it is compared (Diamond 924) with a second constant value which is the low amplitude exponential filter constant UNF EXP .
  • the amplitude of motion is exponentially filtered to yield a filtering value K UNFEXP by setting K UNFEXp (Block 926) to a value equalling the amplitude of motion A DFTUNF divided by the exponential filter constant UNF EXP , quantity squared. Subsequently, as is described below, the amplitude A DFTSNF is multiplied by the exponential filtering value K UNFEXP .
  • the exponential filter constant UNF EXP is 0.5 inch in the described embodiment.
  • the filtering value is set to 1 so that subsequently the effective amplitude of motion remains unchanged (Block 928).
  • the value K UNFEXP and thus the effective amplitude of motion, is set (Block 922) to zero if the amplitude of motion is less than the first constant value.
  • the resulting filtered amplitude of motion (A DFTU NF K UNFEX ) is then multiplied by a vehicle dependent gain factor GAIN UNF , 3.0 by way of example in the described embodiment, to determine the desired rebound damping force (Block 930).
  • the gain factor is 1/4 (1/2 of the constant value of 0.5 inches results in a filter gain factor or exponential filter value of 1/4.
  • the compression force is optional, however, and in the described embodiment is zero.
  • Conventional dampers typically divide the damping into 1/3 for compression and 2/3 for rebound.
  • U.S. Patent 4,634,142 discloses DFT used to detect the UNF frequency and then apply a compression force proportional to the DFT output. Later testing on a vehicle and conventional damper approaches showed that the force should be applied to rebound only and not compression. Therefore, the improved algorithm of FIG. 9 was developed that uses rebound only (Block 930) and also adds small amplitude clipping and exponential filtering to try and help reduce harshness as the UNF motion is damped as in FIG. 9c. Note, the novel use of modulating the force with the output of the DFT remains.
  • FIG. 16 includes: 1) the modulation of the force by the damper velocity (Block 1620), 2) the addition of a maximum limit on the force applied (Block 1625), 3) a falling edge integrator to hold the damping force up as the oscillations damp out (and the output of the DFT falls) (FIG.
  • the integrated position P INTPD forms an average value over a selected period of time.
  • the integration provides an effective low pass filter whereby faster changing signals are attenuated over slower changing signals.
  • the choice of integration constant ⁇ PD and cycle constant N PD determines at what frequency the attenuation starts to significantly take effect. This concept is well known in the art.
  • U.S. Patent No. 4,634,142 discloses that wheel position is integrated and if it is pumping down, additional force is applied to compression to balance the damping and limit the pumping down.
  • the same basic approach but, rather than adding forces to balance the pumping down forces and thus adding more harshness, the preferred approach disclosed herein is to subtract from the rebound forces that are pumping the suspension down (see Blocks 1032, 1034) rather than adding to the compression forces.
  • FIGS, 11a-e show a computational process for determining controller roll response (Block 532).
  • Each cycle intermediate compression and rebound values are calculated using steering angle ⁇ and vehicle speed S.
  • a roll variable R representing the force inducing a roll motion, is calculated by multiplying ⁇ and S.
  • the roll inducement R is determined by retrieving the vehicle speed S and steering angle 0 and multiplying the retrieved values (Block 1112).
  • Intermediate compression and rebound values R C and R R are computed by multiplying the roll inducement value R by vehicle dependent gain compression and rebound factors K RC and K RR , respectively (Block 1114) which in the described embodiment have values of 2.4.
  • a speed dependent roll constant K SR is calculated by retrieving the vehicle speed S and comparing the speed to a minimum value S MIN (Diamond 1120). If S is greater than S MIN , K SR is set to (S-S MIN ) ⁇ (S REFRP -S MIN ) (Block 1122). Otherwise, K SR is set to 0 (Block 1124).
  • S REFRP is 65 miles per hour and S MIN is 30 miles per hour.
  • a differential position R ACT is determined by retrieving the position P LEFT of a left wneel extensible member and the position P RIGHT of an opposite right wheel extensible member and subtracting P RTGHT from P LEFT (Block 1120).
  • the resulting value, R Ac ⁇ is integrated (Block 1132), as has been previously described for other modes, yielding the integrated roll variable R INT .
  • the integration constant ⁇ R used in the described embodiment is 5. This integrated roll variable is then exponentially filtered in a manner similar to that of the unsprung natural frequency amplitude to obtain K REXP .
  • K REXP is set to the magnitude of
  • the desired rebound and compression forces are next computed as shown in FIG. 11d. These computations have been experimentally determined to provide improved ride and handling characteristics.
  • the values K SR , K REXP , R C and R R are retrieved (not shown).
  • a damping compression value COMP is calculated (Block 1142) by adding an optional roll- damping constant R DAMP' which is 10 in the described embodiment, to the intermediate compression value R C , and by multiplying that sum by the filtering roll value K REXP , the speed-dependent roll gain factor K SR and a vehicle dependent gain factor GAIN RP .
  • a damping rebound value REB is similarly calculated using the intermediate roll value R R in place of R C .
  • the desired force to counteract the roll inducement COMPP is calculated (Block 1144) by doubling COMP and adding the resulting value to the product of the intermediate compression value R C and the speed dependent roll gain factor K SR .
  • COMP is approximately 1/3 of COMPP and is applied in the opposite direction to that of COMPP to damp any overshoot as the vehicle attempts to correct itself.
  • the constant value R DAMP that is added provides roll and pitch control which is independent of any sensor indications that control is required. For example, when a vehicle is going straight ahead with a steering angle equalling zero, cross-winds causing roll body motions can be reduced by the position dependent control feature of the algorithm.
  • a desired compression force to counteract the roll motion REBB is calculated in the same manner as COMPP, but using R R in place of R C .
  • the forces COMP, COMPP, REB and REBB are applied to the respective left and right dampers depending upon the direction of induced roll and the direction of the integrated indication of actual roll (using its sign) as shown in FIG. 1 le. Specifically, the values R and R lNT are retrieved (not shown) and if R is greater than or equal to 0 (Block 1152) (indicating that the vehicle is making a right turn), R INT is compared to 0 (Diamond 1154) (indicating the vehicle is rolling left).
  • the above forces are determined whenever the vehicle encounters a maneuver which causes roll or pitch, such as, cornering or braking. Counteracting forces are applied proportional to the amount of influence being applied to the vehicle, such as, the product of vehicle speed and turn angle. In addition, if any motion still occurs (indicating there is incorrect force), then the integrated actual indication of roll R INT provides an error signal so that additional corrective forces are applied or removed in a stabilized feedback manner.
  • the pitch response is determined in a similar manner with the following substitutions.
  • the roll value R is replaced with a pitch value PH equal to the brake pressure p. This represents pitching forward as the vehicle brakes. Acceleration causes the vehicle to squat and any suitable sensor which indicates acceleration can be used as the equivalent of sensing positive brake pressure in the algorithm.
  • the integrated roll variable is replaced with an integrated pitch variable, which utilizes front and rear oppositely placed wheels for extensible member positions, that is, right front and right rear wheels instead of right front and left front wheels. Using these substitutions results in the pitch response desired compression and rebound values. An alternate diagonal comparison is also possible.
  • FIGS. 12a-12e Block diagrams of the pitch response are shown in FIGS. 12a-12e which follow the flow of the roll response of FIGS, 11a-11e.
  • FIGS. 12a-e feature the various "R" subscripts replaced by "PH” subscripts to indicate the change from roll to pitch.
  • Block 1252 determines whether PH ⁇ 0 rather than the R ⁇ 0 determination of Block 1152.
  • the use of a positive only brake pressure to indicate dive (with no indication of squat) precludes certain possibilities shown within FIGS. 12a-e. However, a full-flow diagram is shown to illustrate one possible use with a bi-directional longitudinal acceleration parameter.
  • the pitch response is calculated by separately calculating front and back responses for the left and right sides of the vehicle.
  • the other responses are typically computed serially for each of the four wheels.
  • U.S. Patent No. 4,634,142 discloses that the difference between integrated (low pass filtered) positions on opposite sides/ends are monitored to detect roll/pitch and opposing forces are applied to counteract body motions due to cornering or braking. It has been found that this works on reasonably flat roads, but bumpy roads caused poor response since the filter detection of roll and pitch becomes slower and less accurate. As disclosed above, forces/conditions are detected that would indicate that roll or pitch is about to occur before it does. Added optional sensors for this purpose by way of example these include roll and pitch accelerometers or alternatively a brake sensor, steering and speed sensors as disclosed in FIGS. 7a, e, f and 11a. This improved algorithm provides the correct counteracting forces before the motions even have a chance to occur.
  • the integrated positions disclosed in the '142 patent now act as a feedback stabilizing loops to adjust the forces if and when motions try to occur as in FIGS, 11d, e. Additional algorithm improvements include: 1) modulation of the forces by vehicle speed FIG. 11b, 2) pre-filtering of the position to eliminate high frequency inputs to the DFT (Block 1132), and 3) small amplitude clipping and exponential filtering to try and help reduce harshness (rest of FIG. 11e). Simmlar comments apply to the pitch control of FIG. 12.
  • Bottoming and topping out response is provided to resist the tendency of a vehicle to bottom or top out.
  • the need to provide forces to counter bottoming and topping out is related both to the displacement of the suspension from its equilibrium position and the velocity with which the suspension is travelling farther away from that position.
  • the control system of the present invention provides for the addition of preventative compression and rebound forces to respectively resist bottoming and topping out once a suspension damper has travelled beyond a central zone or window approximately centered about the equilibrium position P LTI .
  • the damper extension corresponding to the upper or compression boundary of the central zone is identified as P BOT (2 inches in the particular embodiment).
  • the damper extension corresponding to the lower or rebound boundary is P TOP (-1-75 inches in the particular embodiment).
  • K PB/T and K VB/T are, respectively, damper position and velocity force constants which, in the described embodiment have values of 8 and 24.
  • differing constants can be provided for bottoming and topping out, respectively.
  • the damper is determined to be in compression or rebound (Diamond 1312). If damper velocity V P is greater than or equal to 0, the suspension is deemed under compression and, if less than 0, under rebound. If the suspension is in compression, it is next determined whether the damper is contracted beyond the upper boundary P BOT of the window (Diamond 1314) by determining whether P is greater than P BOT . if so, the suspension is viewed as about to bottom out and a compression force F CB/T is provided which is equal to BOT and an impulse force F I (the purpose of which is described in detail hereinbelow) is also set to BOT while a rebound force F RB/T is set to zero. (Block 1316).
  • F RB/T is set to F IINT while F CB/T and F I are set to 0 (Block 1330). If F IINT is not greater than 0 and diere is thus no need to counter a prior impulse, F CB/T , F RB/T an F I are all set to 0 (Block 1340).
  • F I is compared to F INTn-1 (Diamond 1350). If greater, F IINT is increased by a constant ⁇ I (Block 1352) and if less, F IINT is decreased by the constant ⁇ I (Block 1354).
  • the constant ⁇ I is equal to 5.
  • the aforementioned pitch and roll responses are provided to resist pitch and roll moments due primarily to respective longitudinal and lateral accelerations. Pitch and roll moments can also be introduced when encountering a bump. For example, if the left front wheel of the vehicle hits a bump, there is a tendency for the vehicle to have a combined squat and roll right movement. This may be approximated as a rotation about the diagonal axis between the right front and left rear wheels compressing the right rear suspension.
  • the stored energy response compensates by having the right rear damper apply a compression- resisting force.
  • the mode associated with such movement is identified as me Stored Energy mode. This mode is especially important where the bump or other influence causes displacement at a frequency near the sprung natural frequency of the vehicle, as higher frequency bumps do not cause so significant a rotation.
  • compression and rebound stored energy responses F CSE and F RSE are determined based upon the position and velocity of both that wheel (the wheel under control or WUC) and the wheel diagonally opposite (diagonal wheel or DW).
  • U.S. Patent No. 4,634,142 discloses that the bottoming and topping forces are a function of how close and how fast the suspension is to reaching its limit, a major improvement in the damping of the resulting motions that occur. More specifically, the ' 142 patent discusses that when a big bump is encountered and forces are applied to avoid hitting the suspension limit (bottom or top), the mass is set into motion and the response is improved if a portion of the energy injected into the system is removed.
  • the present invention is an improvement to the prior method in that it measures how much energy is injected during a bump and then remove it (see Blocks 1350, 1352, 1345 and diamonds 1318, 1328 with their connected blocks).
  • P is subject to a short-term integration (Block 1410) to yield an integrated value P INTSE for the current WUC.
  • the integration constant ⁇ SE is 3 in the described embodiment. The details of the integration and the counter are not shown, but follow those described with regard to the previous modes.
  • P INTSE is clipped and subject to an exponential filter. In particular, if me absolute value of P JNTSE ⁇ s less tl ⁇ an a clipping value SE CLIP (0.5 inch in the particular implementation) (Diamond 1412) an intermediate stored energy parameter K SEEXP is set to 0 (Block 1414).
  • K SEEXP is set to 1 (Block 1418). Otherwise, K SEEXP is set to P INTSE divided by SE EXP , quantity squared (Block 1420).
  • FIG. 14a The process of FIG. 14a is repeated for the diagonal wheel (DW) so as to yield preliminary compression and rebound forces F DWC and F DWR , respectively.
  • the preliminary compression rebound forces for the WUC and DW are selectively combined to yield the final compression and rebound forces in the stored energy mode, F CSE and F RSE , respectively. If the integrated position P INTSE of the WUC and DW are wittiin a window defined by limits ⁇ SE LIM (1 inch in the particular implementation), no compression or rebound forces are provided.
  • F CSE is set to 0 and F RSE is set to the absolute value of F DWR minus the absolute value of F WUCR (Block 1438). If the calculated F RSE would be less dian (Diamond 1439), it is set to 0 (Block 1440).
  • U.S. Patent No. 4,634,142 discloses that the stored energy control applies counteracting forces to control the body motions.
  • the detailed algorithms disclosed herein implement stored energy control.
  • An improvement over the '142 patent is to add small amplitude clipping and exponential filtering to help reduce harshness (see FIG. 14a and Blocks 1430-1433, 1442, 1448).
  • the float response provides a small constant for damper force with respect to position for when the amplitude of motion at the sprung natural frequency or the amount of actual roll and pitch exceed certain minimum values.
  • the constant or "Coulomb" force is dependent upon the vehicle velocity. In the described embodiment, it is applied in the rebound direction only as rebound damping forces are less perceptible than compression damping forces from the point of view of a vehicle passenger. Force is applied only to the extent that the other response modes are not providing such force.
  • each cycle to determine whether there is motion near the sprung natural frequency, the current and previous values of the integrated position P SNF and P SNFn-1 are received and compared (Diamond 1512) for determining whether the damper is under rebound or compression. If P SNF is greater than or equal to P SNFn-1 , damper is not in rebound and therefore it is meaningless to apply a rebound resisting force. Thus, P RFL is set to 0 (Block 1514) (F CFL is always 0).
  • the integrated sprung natural frequency parameter A INTSNF is retrieved (not shown) and compared to a threshold float amplitude FL SNFLIM which is the minimum sprung natural frequency movement for which Coulomb damping is applied (0.25 inches in the described embodiment) (Diamond 1518). If A INTSNF is less than FL SNFLIM' then no Coulomb force is necessary to damp sprung natural frequency motion and, in similar fashion, it is determined whether Coulomb force is required to damp roll motion by retrieving the integrated actual roll R INT (not shown) and comparing it to a threshold float amplitude FL RLIM (0.25 inches in the described embodiment) (Diamond 1522).
  • R INT is less man FL RLI M
  • no Coulomb force is necessary to damp roll motion and it is similarly determined whether a Coulomb force is necessary to damp pitch motion by retrieving the integrated actual pitch PH INT (not shown) and comparing that value to a threshold float amplitude FL PLIM (0.25 inches in the described embodiment) (Diamond 1526). If PH INT is less than FL PLIM , there is no need for a Coulomb force and F RFL is set to 0 (Block 1514).
  • a necessary Coulomb force COUL is determined by retrieving the vehicle velocity dependent gain factor K SR (not shown) and multiplying it by a constant force FL COUL (200 lbs. in the described embodiment) (Block 1530).
  • COUL is men compared to the sum of the rebound forces due to the sprung natural frequency, unsprung natural frequency, roll and pitch, pumping down, bottoming and topping out, and stored energy modes (Diamond 1532). If COUL is less dian this sum, there is no need for additional float damping and F RFL is set to 0 (Block 1514). If COUL exceeds this sum, there is need for additional damping and F RFL is set to the difference between COUL and the sum (Block 1534) so that when summed, the net rebound force F R will equal COUL.
  • lateral (roll) accelerometer 60 and longitudinal (pitch) accelerometer 61 can be used as an alternative or a supplement as shown in FIG. 18.
  • a longitudinal accelerometer can be utilized within the pitch mode to provide anti-squat forces upon acceleration in addition to the anti-dive forces provided upon braking.
  • the longitudinal acceleration can be obtained by differentiating the output of the speed sensor.
  • brake pressure engine vacuum pressure can be used to indicate acceleration.
  • FIGS. 19 and 20 block diagrams of roll control response calculations and pitch control response calculations, respectively, are shown for use with the alternate computer-controlled system of FIG. 18.
  • the block diagram of FIG. 19 corresponds to the block diagram of FIG. 11a, with the exception of roll acceleration factor R A in place of vehicle speed S and steering angle ⁇ (Block 1912).
  • vehicle-dependent gain factors K RAC and K RAR are used (Block 1914).
  • the block diagram of FIG. 20 corresponds to the block diagram of FIG. 12a.
  • the brake pressure ⁇ is replaced by pitch acceleration factor P A (Block 2012).
  • the vehicle-dependent gain factors K PHAC and K PHAR are used (Block 2014).

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Automation & Control Theory (AREA)
  • Vehicle Body Suspensions (AREA)

Abstract

La présente invention concerne un contrôleur (16) destiné à la gestion d'un système d'amortisseurs (12). Ce système (12) se compose d'au moins deux amortisseurs (12a - 12d) assurant l'amortissement entre des masses suspendues par ressorts et non suspendues par ressorts (7, 2) dans le sens de la compression et/ou du rebond. Un capteur (190a - 190d) génère des signaux de position (17a - 17d) représentant le déplacement entre les masses suspendues par ressorts et non suspendues par ressorts (7, 2). Un régulateur (40a, 40b) réagit à l'un au moins des signaux de commande de compression et de rebond indépendants (25a - 25d, 27a - 27d), afin de réguler, respectivement, l'une au moins des forces de résistance à la compression et au rebond des amortisseurs (12a - 12d) entre les masses suspendues par ressorts et non suspendues par ressorts (7, 2). Le contrôleur (16) comporte un processeur (15) qui réagit aux signaux représentant les signaux de positions (17a - 17d), ce qui permet de produire les signaux de commande de compression et de rebond (25a - 25d, 27a - 27d) destinés au régulateur (40a, 40b), en fonction du mouvement entre les masses (2, 7) ou du mouvement d'un véhicule (8) dans lequel se trouvent les amortisseurs (12a - 12d).
PCT/US1995/010641 1994-08-18 1995-08-18 Systeme de suspension adaptative optimise par ordinateur et ameliorations apportees au procede associe Ceased WO1996005975A1 (fr)

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US08/293,539 1994-08-18

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EP0921387A3 (fr) * 1997-12-08 1999-12-08 Beissbarth GmbH Méthode et dispositif pour tester des amortisseurs montés
EP0921386A3 (fr) * 1997-12-08 2000-01-19 Beissbarth GmbH Méthode et dispositif pour tester des amortisseurs montés
WO2000027658A1 (fr) 1998-11-11 2000-05-18 Kenmar Company Trust Unite de commande de suspension et vanne de commande
US10987989B2 (en) 2017-06-09 2021-04-27 Polaris Industries Inc. Adjustable vehicle suspension system
US10987987B2 (en) 2018-11-21 2021-04-27 Polaris Industries Inc. Vehicle having adjustable compression and rebound damping
US11110913B2 (en) 2016-11-18 2021-09-07 Polaris Industries Inc. Vehicle having adjustable suspension
US11124036B2 (en) 2012-11-07 2021-09-21 Polaris Industries Inc. Vehicle having suspension with continuous damping control
US11285964B2 (en) 2014-10-31 2022-03-29 Polaris Industries Inc. System and method for controlling a vehicle
US11332071B2 (en) 2004-10-05 2022-05-17 Vision Works Ip Corporation Absolute acceleration sensor for use within moving vehicles
US11407357B2 (en) 2013-08-28 2022-08-09 Vision Works Ip Corporation Absolute acceleration sensor for use within moving vehicles
US20220297495A1 (en) * 2021-03-16 2022-09-22 Ford Global Technologies, Llc Suspension system with jump control and/or whoop detection
US11577705B2 (en) 2004-10-05 2023-02-14 VisionWorks IP Corporation Absolute acceleration sensor for use within moving vehicles
US11904648B2 (en) 2020-07-17 2024-02-20 Polaris Industries Inc. Adjustable suspensions and vehicle operation for off-road recreational vehicles
US11963064B2 (en) 2016-02-10 2024-04-16 Polaris Industries Inc. Recreational vehicle group management system
US12038301B2 (en) 2013-02-26 2024-07-16 Polaris Industries Inc. Recreational vehicle interactive telemetry, mapping and trip planning system
US12163799B2 (en) 2013-02-26 2024-12-10 Polaris Industries Inc. Recreational vehicle interactive telemetry, mapping, and trip planning system
US12297785B2 (en) 2004-10-05 2025-05-13 Vision Works Ip Corporation Absolute acceleration sensor for use within moving vehicles
US12391116B2 (en) 2010-06-03 2025-08-19 Polaris Industries Inc. Adjustable performance for a vehicle
US12397878B2 (en) 2020-05-20 2025-08-26 Polaris Industries Inc. Systems and methods of adjustable suspensions for off-road recreational vehicles

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EP0921387A3 (fr) * 1997-12-08 1999-12-08 Beissbarth GmbH Méthode et dispositif pour tester des amortisseurs montés
EP0921386A3 (fr) * 1997-12-08 2000-01-19 Beissbarth GmbH Méthode et dispositif pour tester des amortisseurs montés
AU735047B2 (en) * 1997-12-08 2001-06-28 Beissbarth Gmbh Shock absorber tester II
WO2000027658A1 (fr) 1998-11-11 2000-05-18 Kenmar Company Trust Unite de commande de suspension et vanne de commande
WO2000028236A1 (fr) 1998-11-11 2000-05-18 Kenmar Company Trust Systeme adaptatif de suspension optimise par ordinateur et methode correspondante
US6296091B1 (en) 1998-11-11 2001-10-02 Kenmar Company Trust Suspension control unit and control valve
US6502837B1 (en) 1998-11-11 2003-01-07 Kenmar Company Trust Enhanced computer optimized adaptive suspension system and method
US11332071B2 (en) 2004-10-05 2022-05-17 Vision Works Ip Corporation Absolute acceleration sensor for use within moving vehicles
US12297785B2 (en) 2004-10-05 2025-05-13 Vision Works Ip Corporation Absolute acceleration sensor for use within moving vehicles
US11577705B2 (en) 2004-10-05 2023-02-14 VisionWorks IP Corporation Absolute acceleration sensor for use within moving vehicles
US12391116B2 (en) 2010-06-03 2025-08-19 Polaris Industries Inc. Adjustable performance for a vehicle
US11400786B2 (en) 2012-11-07 2022-08-02 Polaris Industries Inc. Vehicle having suspension with continuous damping control
US12291069B2 (en) 2012-11-07 2025-05-06 Polaris Industries Inc. Vehicle having suspension with continuous damping control
US11400787B2 (en) 2012-11-07 2022-08-02 Polaris Industries Inc. Vehicle having suspension with continuous damping control
US11400784B2 (en) 2012-11-07 2022-08-02 Polaris Industries Inc. Vehicle having suspension with continuous damping control
US11400785B2 (en) 2012-11-07 2022-08-02 Polaris Industries Inc. Vehicle having suspension with continuous damping control
US11970036B2 (en) 2012-11-07 2024-04-30 Polaris Industries Inc. Vehicle having suspension with continuous damping control
US11124036B2 (en) 2012-11-07 2021-09-21 Polaris Industries Inc. Vehicle having suspension with continuous damping control
US12163799B2 (en) 2013-02-26 2024-12-10 Polaris Industries Inc. Recreational vehicle interactive telemetry, mapping, and trip planning system
US12038301B2 (en) 2013-02-26 2024-07-16 Polaris Industries Inc. Recreational vehicle interactive telemetry, mapping and trip planning system
US11407357B2 (en) 2013-08-28 2022-08-09 Vision Works Ip Corporation Absolute acceleration sensor for use within moving vehicles
US11285964B2 (en) 2014-10-31 2022-03-29 Polaris Industries Inc. System and method for controlling a vehicle
US11919524B2 (en) 2014-10-31 2024-03-05 Polaris Industries Inc. System and method for controlling a vehicle
US12325432B2 (en) 2014-10-31 2025-06-10 Polaris Industries Inc. System and method for controlling a vehicle
US11963064B2 (en) 2016-02-10 2024-04-16 Polaris Industries Inc. Recreational vehicle group management system
US11110913B2 (en) 2016-11-18 2021-09-07 Polaris Industries Inc. Vehicle having adjustable suspension
US12337824B2 (en) 2016-11-18 2025-06-24 Polaris Industries Inc. Vehicle having adjustable suspension
US11878678B2 (en) 2016-11-18 2024-01-23 Polaris Industries Inc. Vehicle having adjustable suspension
US12330467B2 (en) 2017-06-09 2025-06-17 Polaris Industries Inc. Adjustable vehicle suspension system
US10987989B2 (en) 2017-06-09 2021-04-27 Polaris Industries Inc. Adjustable vehicle suspension system
US11912096B2 (en) 2017-06-09 2024-02-27 Polaris Industries Inc. Adjustable vehicle suspension system
US11479075B2 (en) 2017-06-09 2022-10-25 Polaris Industries Inc. Adjustable vehicle suspension system
US11884117B2 (en) 2018-11-21 2024-01-30 Polaris Industries Inc. Vehicle having adjustable compression and rebound damping
US10987987B2 (en) 2018-11-21 2021-04-27 Polaris Industries Inc. Vehicle having adjustable compression and rebound damping
US11975584B2 (en) 2018-11-21 2024-05-07 Polaris Industries Inc. Vehicle having adjustable compression and rebound damping
US12384214B2 (en) 2018-11-21 2025-08-12 Polaris Industries Inc. Vehicle having adjustable compression and rebound damping
US12397878B2 (en) 2020-05-20 2025-08-26 Polaris Industries Inc. Systems and methods of adjustable suspensions for off-road recreational vehicles
US11904648B2 (en) 2020-07-17 2024-02-20 Polaris Industries Inc. Adjustable suspensions and vehicle operation for off-road recreational vehicles
US20220297495A1 (en) * 2021-03-16 2022-09-22 Ford Global Technologies, Llc Suspension system with jump control and/or whoop detection

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