U.S. patent application number 10/144625 was filed with the patent office on 2002-10-24 for system for control of active system for vibration and noise reduction.
Invention is credited to Kiss, John C., Silverberg, Michael H..
Application Number | 20020153451 10/144625 |
Document ID | / |
Family ID | 25173357 |
Filed Date | 2002-10-24 |
United States Patent
Application |
20020153451 |
Kind Code |
A1 |
Kiss, John C. ; et
al. |
October 24, 2002 |
System for control of active system for vibration and noise
reduction
Abstract
An adaptive controller is used to adaptively generate vibration
cancellation signals driving a controlled device which effects an
associated vibration and noise-producing plant. The adaptive
controller has multiple control paths to generate the control
signal. In a vibration attenuation control path(s), an adaptive
control signal is generated by plant compensation and adaptive
filtering techniques to cancel vibrations. In a position control,
saturation prevention path, the available operational extents of
the controlled device are monitored and compensation signals are
generated which direct the movement of the controlled device in
such a manner as to prevent the controlled device from reaching the
extents of control. The control signals from the multiple paths are
then combined and transmitted to the controlled device which alters
in some fashion the noise and vibration being generated or
transmitted by the associated vibrating plant.
Inventors: |
Kiss, John C.; (Edison,
NJ) ; Silverberg, Michael H.; (Livingston,
NJ) |
Correspondence
Address: |
Michael G. Johnston
Moore & Van Allen
Suite 800
2200 West Main Street
Durham
NC
27705
US
|
Family ID: |
25173357 |
Appl. No.: |
10/144625 |
Filed: |
May 13, 2002 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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10144625 |
May 13, 2002 |
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09798420 |
Mar 2, 2001 |
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6402089 |
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Current U.S.
Class: |
244/17.27 |
Current CPC
Class: |
B64C 2220/00 20130101;
B64C 27/001 20130101 |
Class at
Publication: |
244/17.27 |
International
Class: |
B64C 027/00 |
Goverment Interests
[0001] The Government has rights to the invention pursuant to
government contract N00014-96-C-2079 awarded by the United States
Naval Research Laboratory.
Claims
I claim:
1. A control system for reducing vibration generated by a vibrating
plant including a vibrating component, a structure and a mount for
mounting the vibrating component to the structure, the control
system comprising: a. means for producing controlled vibrations
within the mount; b. at least one sensor for sensing the current
position of the controlled vibration producing means and developing
a signal indicative thereof; c. at least one sensor for sensing the
vibration being transmitted through the mount from the vibrating
component to the structure and developing a signal indicative
thereof; d. at least one sensor for sensing at least one of the
characteristic frequencies at which the vibrating plant operates
and developing a signal indicative thereof; e. a first controller
receiving as input the position signal developed by the controlled
vibration producing means sensor and generating an output signal;
f. a second controller receiving as input the transmitted vibration
sensor signal and the characteristic sensor signal and generating
an output signal; and g. means for combining the output signals
from the first and second controllers into a control signal for
controlling the vibration producing means such that the vibration
transmitted from the vibrating component to the structure through
the mount is reduced.
2. The control system according to claim 1, wherein the first
controller responds to the controlled vibration producing means
position sensor by providing output signals which maintain a
desired position of the controlled vibration producing means.
3. The control system according to claim 1, wherein the first
controller produces a quasi-static, fixed, low bandwidth, broadband
control signal.
4. The control system according to claim 3, wherein the first
controller comprises a compensation function utilizing
proportional, integral, derivative control.
5. The control system according to claim 1, wherein the second
controller produces a narrow-band feedback vibration control
signal.
6. The control system according to claim 5, wherein the second
controller comprises a. a frequency filter adaptive to isolate
sensed vibration signals at frequencies which are multiples of the
sensed characteristic plant frequency; b. an objective function
characterizing the magnitude of said isolated signals; c. a
compensation function producing a correlation between said isolated
signals and said control signal for said controlled vibration
producing means; and d. an adaptive filter which generates
attenuation output signals minimizing said isolated, correlated
signals.
7. An active vibration control system for reducing vibration
generated by a vibrating plant including a vibrating component, a
structure and a hydraulic mount for mounting the vibrating
component to the structure, the vibration transmitted from the
vibrating component through the hydraulic mount to the structure,
the control system comprising: a. at least one hydraulic actuator
for producing controlled vibrations within the mount; b. at least
one position sensor for sensing the current position of the
hydraulic actuator and producing a signal representative thereof;
c. at least one vibration sensor for sensing vibrations being
transmitted from the vibrating component through the hydraulic
mount to the structure and producing a signal representative
thereof; d. at least one vibrating plant sensor for sensing at
least one of the rotational frequencies at which said vibrating
plant operates and producing a signal representative thereof; e. a
fixed, low bandwidth, near-DC, PID-based broadband control
compensation feedback position controller which utilizes the
position sensor signal to produce position control signals to
minimize the offset between the sensed hydraulic actuator position
and a predetermined hydraulic actuator position; f. an adaptive
Filtered-X LMS based narrow-band vibration controller which
utilizes the vibration sensor signal to produce vibration control
signals at multiple frequencies of the sensed plant characteristic
frequencies; and g. means for combining position control signals
with the vibration control signals and generating an output signal,
the hydraulic actuator being responsive to the output signal of the
combining and output signal generating means for producing
controlled vibrations in the mount for reducing vibrations
transmitted through the mount from the vibrating component to the
structure.
8. The control system according to claim 7, wherein the position
controller comprises: a. a scaling function; b. a band elimination
function; c. an objective function; and d. a compensation
function.
9. The control system according to claim 8, wherein the
compensation function produces the position control signal
utilizing proportional, integral, derivative control
compensation.
10. The control system according to claim 7, wherein the vibration
controller comprises: a. a frequency filter adaptive to isolate
sensed vibration signals at frequencies which are multiples of the
sensed plant characteristic frequency; b. an objective function
characterizing the magnitude of the isolated signals; c. a
compensation function producing a correlation between the isolated
signals and the control signal for the controlled vibration
producing means; and d. an adaptive filter which generates
attenuation output signals minimizing the isolated, correlated
signals.
11. The control system according to claim 10, wherein the vibration
controller further comprises: a. a frequency downshift function
which converts the vibration sensor signals to signals at baseband
DC; and b. a frequency upshift function which converts the baseband
DC signals into in-band, attenuation path-based, vibration control
signals.
12. The control system according to claim 11, wherein the vibration
controller further comprises: a. an input function which performs
antialiasing and scaling functions on the vibration sensor signals;
b. a normalization function which normalizes the isolated signals;
c. an output function which scales the vibration control signals;
and d. a weight limiting function which evaluates the vibration
control signals and transmits a freeze signal to the adaptive
filter function affecting the adaptive abilities of the adaptive
filter function.
13. The control system according to claim 12, wherein the frequency
filter comprises: a. a band-pass filter; and b. a notch filter
receiving as input the plant rotational frequency, the notch filter
adapting its filter window based on the input plant rotational
frequency.
14. An active vibration control system for reducing vibration in a
rotary wing aircraft, the rotary wing aircraft including an
airframe and a main rotor system having a rotor, an engine and a
transmission gearbox mounted to the airframe for turning the engine
force into the rotational force of a rotorshaft, the gearbox
attached to the airframe with at least one hydraulic mount between
the gearbox and the airframe, wherein the operation of the main
rotor system generates vibration that is transferable to the
airframe causing vibration on board the aircraft, , the control
system comprising: a. at least one hydraulic actuator for producing
controlled vibrations within the mount; b. at least one position
sensor for sensing the current position of the hydraulic actuator
and producing a signal representative thereof; c. at least one
vibration sensor for sensing vibrations being transmitted from the
main rotor system through the mount to the airframe and producing a
signal representative thereof; d. at least one rotational sensor
for sensing the rotational frequency of the rotorshaft and
producing a signal representative thereof; e. a fixed, low
bandwidth, near-DC, broadband control compensation feedback
position controller which utilizes the position sensor signal to
produce quasi-static position control signals to minimize the
offset between the sensed hydraulic actuator position and a
predetermined hydraulic actuator position; f. an adaptive
Filtered-X LMS based narrow-band vibration controller which
utilizes the vibration sensor signal to produce vibration control
signals ; and g. means for combining the position control signals
with the vibration control signals and generating an output signal,
the hydraulic actuator being responsive to the output signal of the
combining and output signal generating means for producing
controlled vibrations in the mount for reducing vibrations
transmitted through the mount from the engine and rotor system to
the airframe.
15. The control system according to claim 14, wherein the position
controller comprises: a. a scaling function; b. a band elimination
function; c. an objective function; and d. a compensation
function.
16. The control system according to claim 15, wherein the
compensation function produces the position control signal
utilizing proportional, integral, derivative control
compensation.
17. The control system according to claim 14, wherein said
vibration controller comprises: a. a frequency filter adaptive to
isolate sensed vibration signals at frequencies which are multiples
of the sensed rotorshaft frequency; b. an objective function
characterizing the magnitude of the isolated signals; c. a
compensation function producing a correlation between the isolated
signals and the control signal for the; and d. an adaptive filter
which generates attenuation output signals minimizing the isolated,
correlated signals.
18. The control system according to claim 17, wherein the vibration
controller further comprises: a. a frequency downshift function
which converts the vibration sensor signals to signals at baseband
DC; and b. a frequency upshift function which converts the baseband
DC signals into in-band, attenuation path-based, vibration control
signals.
19. The control system according to claim 18, wherein said
vibration controller further comprises: a. an input function which
performs antialiasing and scaling functions on the vibration sensor
signals; b. a normalization function which normalizes the isolated
signals; c. an output function which scales the vibration control
signals; and d. a weight limiting function which evaluates the
vibration control signals and transmits a freeze signal to the
adaptive filter function affecting the adaptive abilities of the
adaptive filter function.
20. The control system according to claim 19, wherein the frequency
filter comprises: a. a band-pass filter; and b. a notch filter
receiving as input the rotorshaft rotation frequency, the notch
filter adapting its filter window based on the input rotorshaft
rotation frequency.
21. A method of controlling vibration suppression apparatus means
in a mechanical system, the method comprising the steps of: sensing
a position state of at least one vibrating component in the
mechanical system to produce a position sensor signal; sensing
vibrations of the at least one vibrating component in the
mechanical system to produce a vibration sensor signal; applying a
position control function to the position sensor signal to produce
a position control function output signal; applying a vibration
control function to the vibration sensor signal to produce a
vibration control function output signal; and combining the
position control function output signal with the vibration control
function output signal to produce a control signal that operates
the vibration suppression apparatus.
22. The method of claim 21, wherein the position control function
includes proportional broadband control compensation feedback.
23. The method of claim 21, wherein the vibration control function
includes adaptive, filtered-X least-mean-square narrowband
control.
24. The method of claim 22, wherein the vibration control function
includes adaptive, filtered-X least-mean-square narrowband control.
Description
CROSS-REFERENCES
FIELD OF THE INVENTION
[0002] The present invention relates to a control system for an
active, adaptive vibration and noise attenuation system. The
present invention serves as the intelligence of an overall system
that has several parts. Generally, the other parts of the control
system are sensors for measuring the objectionable vibration and
noise and one or more controlled devices for providing a mechanism
for altering the production of noise and vibration. In particular,
the present invention relates to a control system combining the
results of multiple paths to generate a resulting vibration and
noise control signal with at least one attenuation path used to
generate vibration and noise attenuation signals and at least one
other path used to generate signals which seek to control the
position of the altering mechanism to prevent saturation of the
mechanism.
[0003] The present invention also relates generally to a system for
controlling an active system for reducing the transmission of
vibration and noise passing from a vibrating component to a
structure and, more particularly, to a system for controlling an
active vibration and noise reduction system for use on a rotary
wing aircraft.
[0004] Even more particularly, the present invention comprises an
Active Transmission Mount Controller (ATM Controller) to be used to
control a number of hydraulic actuation systems utilized in active
cancellation of vibration in rotary wing aircraft. The ATM
Controller controls hydraulic actuators located in-line between
each transmission foot and the airframe. The ATM Controller
produces outputs that are based on the fundamental blade rotational
rate as well as multiples of this rate. In addition, the ATM
Controller produces a position control signal to maintain the
relative position of the transmission foot and the aircraft.
BACKGROUND OF THE INVENTION
[0005] Significant effort has been devoted to reducing the
vibratory and acoustic loads on aircraft, particularly rotary wing
aircraft such as helicopters, and the resulting vibration and noise
that develops within the aircraft. A primary source of vibratory
and acoustic loads in a helicopter is the main rotor system.
[0006] The main rotor system of a helicopter includes rotor blades
mounted on a vertical shaft that projects from a transmission,
often referred to as a gearbox. The gearbox comprises a number of
gears which reduce the rotational speed of the helicopter's engine
to the much slower rotational speed of the main rotor blades. The
gearbox has a plurality of mounting "feet" which are connected
directly to structure in the airframe which supports the
gearbox.
[0007] The main rotor lift and driving torque produce reaction
forces and moments on the gearbox. All of the lift and maneuvering
loads are passed from the main rotor blades to the airframe through
the mechanical connection between the gearbox feet and the
airframe. The airframe structure which supports the gearbox is
designed to react to these primary flight loads and safely and
efficiently transmit the flight loads to the airframe.
[0008] In addition to the nearly static primary flight loads, the
aircraft is also subjected to vibratory loads originating from the
main rotor blades and acoustic loads generated by clashing of the
main rotor transmission gears. The vibratory loads are strongest at
a frequency equal to the rotational speed of the main rotor blades
(P), which is generally between about 4 and about 5 Hz, multiplied
by the number of rotor blades, typically 2 or 4. The product of the
main rotor blades rotational speed and the number of blades is
called the "fundamental". Tonals of decreasing vibratory strength
occur at multiples of two, three and sometimes four of the
fundamental. For example, for a 4 bladed rotor, this would
correspond to 8P, 12P, and 16P.
[0009] The acoustic loads generated by the transmission gears are
at a frequency that the gear teeth mesh with and contact each
other, and are thus related to the type of construction and gear
ratios used in the transmission. The acoustic loads also include a
fundamental and tonals of decreasing strength at integer multiples
of the fundamental. Typically, the noise generated by gear clashing
is in the range of about 500 Hz to about 3 kHz.
[0010] The vibratory and acoustic loads produce vibrations and
audible noise that are communicated directly to the helicopter
airframe via the mechanical connection between the gearbox and the
airframe. This mechanical connection becomes the "entry point" for
the unwanted vibration and noise energy into the helicopter cabin.
The vibrations and noise within the aircraft cabin cause discomfort
to the passengers and crew. In addition, low frequency rotor
vibrations are a primary cause of maintenance problems in
helicopters.
[0011] In the past, "passive" solutions have been tried for
reducing the vibratory and acoustic loads on aircraft and the
resulting vibration and noise that develops within the aircraft.
For noise reduction, passive systems have employed broadband
devices such as absorbing blankets or rubber mounts. However,
broadband passive systems have generally proven to be heavy and,
consequently, not structurally efficient for aircraft applications
where weight is paramount. Additionally, broadband passive systems
are not very effective at reducing low frequency vibration. A
passive technique for reducing vibration involves the installation
of narrowband, low frequency vibration absorbers around the
aircraft that are tuned to the vibration frequency of interest,
typically the fundamental. These narrowband, passive vibration
reduction systems are effective, but the number of vibration tonals
present in a helicopter may require a number of these systems which
then adds significant weight. Additionally, narrowband passive
systems work best when placed at ideal locations about the
helicopter airframe, many of which may be occupied by other
equipment.
[0012] More recently, "active" vibration and noise reduction
solutions are being employed since active systems have a much lower
weight penalty and can be effective against both low frequency
vibration and higher frequency noise. Active systems utilize
sensors to monitor the status of the aircraft, or the vibration
producing component, and a computer-based controller to command
countermeasures to reduce the vibration and noise. The sensors are
located throughout the aircraft and provide signals to the adaptive
controller. The controller provides signals to a plurality of
actuators that are located at strategic places within the aircraft.
The actuators produce controlled forces or displacements which
attempt to minimize vibration and noise at the sensed
locations.
[0013] Low frequency motion (i.e., vibration) behaves according to
rigid body rules and structural models can be used to accurately
predict the nature and magnitude of the motion. Since low frequency
motion is easily modeled, its negative effects can be cancelled
with an active system of moderate complexity. High frequency motion
(i.e., noise) at the transmission gear clash frequencies does not
obey the rigid body rules present at low vibration frequencies. The
use of riveted airframes in combination with the complex mode
shapes present at high frequencies makes structural models much
less accurate. As a result, active systems for high frequency
energy reduction become more complex, requiring large numbers of
actuators and sensors to counter the more complex modal behavior of
the airframe structure.
[0014] Some active systems utilize hydraulic actuation systems and
hydraulic actuators to reduce vibration and noise. The hydraulic
actuation system is preferred since the hydraulic system provides
the necessary control bandwidth and authority to accommodate the
frequencies and high loads typically experienced in an aircraft
such as a helicopter. Additionally, aircraft typically have
hydraulic power sources with spare capacity which can be utilized
or augmented.
[0015] Two methods of actuator placement are frequently used in
active systems: (1) distribute the actuators over the airframe, or
(2) co-locate the actuators at, or near, the vibration or noise
entry point. The co-location approach places the actuators at or
near the structural interface between the transmission and airframe
stopping vibration and noise near the entry point before it is able
to spread out into the aircraft. This has the advantage of reducing
the number of actuators and the complexity of the control system.
Active systems using this approach employ actuators mounted in
parallel or in series with the entry point to counteract the
vibration and noise.
[0016] The distributed actuator approach requires a large number of
actuators for controlling noise due to the high frequencies, and
their associated short spatial wavelength. The large number of
actuators can drive up weight and add significantly to control
system complexity. One distributed actuator active noise reduction
system for use in a helicopter application uses more than 20
actuators to control transmission noise. Distributed actuators for
low frequency vibration will be less numerous and are effective at
reducing vibration at the sensor locations, but can drive vibration
at other areas of the aircraft to levels exceeding those already
present.
[0017] The parallel actuator approach is effective for low
frequency vibration but can produce counteracting forces in the
supporting structural elements which can exceed the design limit of
the elements and lead to premature failure. Additionally, the
parallel approach is not effective at reducing noise because the
parallel actuators provide a direct path for noise entry.
[0018] The series approach is the most effective in reducing cabin
vibration and avoids the introduction of unwanted vibrations. This
approach uses actuators mounted in series between the transmission
gearbox feet and airframe support structure. In this approach, the
gearbox and airframe are isolated from each other connected only by
actuators. The gearbox vibrates in its own inertial frame
separately from the airframe's inertial frame, isolating the
gearbox and airframe in a dynamic sense. This approach interrupts
the transmission of vibratory and acoustic energy through the
principal entry point thereby preventing vibration and noise from
entering the airframe. For this approach to be effective, the
vibration and noise isolation system must support the large, static
primary flight loads along an axis also requiring dynamic
isolation. This system must also maintain the average static
position of the transmission relative to the airframe for proper
operation of the other helicopter systems, particularly the
helicopter engines that couple into the transmission. However, in
the series approach, the high frequencies associated with noise
lead to complex motions at the entry point which, if fully
addressed, may lead to large and heavy actuators to actively
control all degrees of freedom at each entry point.
[0019] A more efficient way for reducing both vibration and noise
in aircraft applications, and particularly helicopters, combines an
active system for low frequency vibration reduction with a passive
system for high frequency noise reduction. Preferably, the active
vibration reduction system will isolate the vibratory load source,
such as the main rotor system of the helicopter, and prevent the
low frequency vibration generated by the main rotor system from
being transmitted to the airframe. The system should efficiently
pass the primary flight loads while maintaining the average static
position of the gearbox relative to the airframe.
[0020] Adaptive controllers for active vibration reduction systems
are well known in the art. These controllers monitor vibrations and
seek to generate signals which drive devices producing canceling
vibrations. The controlled devices used to cancel vibrations act
either upon the body producing the objectionable vibrations or the
controlled devices may act upon some connection point between the
vibration generating machinery and the vibration measurement point.
Such connection point efforts include actuators which connect
helicopter transmission feet to helicopter cabins.
[0021] One method known in the art is to measure the noise and
vibration disturbances at locations where cancellation is desired
and to feedback this information into an active controller which
then makes alteration/cancellation adjustments to reduce the noise
and vibration disturbances. Feedback systems tend to be effective
when the time delay through the controller actuator and sensors is
kept to a minimum.
[0022] Existing adaptive controllers assume sufficient authority
exists in the vibration cancellation mechanism to respond to the
vibration cancellation signals. This may not always be true. For
example, a hydraulic actuator used to produce cancellation
vibrations may reach the maximum extent of actuation. In such a
situation, the actuator could not continue to respond to
cancellation signals until the actuator moves sufficiently away
from a maximum actuation extent. Cessation of ability to respond
has at least two drawbacks. The first is an obvious reduction in
the cancellation of the vibration being controlled by the impaired
cancellation mechanism. The second drawback is that a mechanism
such as an actuator at full extent may exhibit characteristics
similar to a fixed mount. Such a fixed mount might reduce the
effectiveness of passive vibration reduction techniques used in
conjunction with the active vibration control system.
[0023] For the foregoing reasons, there is a need for a new control
system for active reduction of both vibration and noise. The new
controller will transmit output vibration cancellation signals
which control an active vibration cancellation mechanism. Such
vibration cancellation mechanism will be located within the
connection points and in series between a vibration generating
component and the mounting location of the component. The
controller should employ two or more control paths to ensure that
the vibration cancellation mechanism maintains the relative
position between the vibration generating component and the
mounting location and has sufficient authority to respond to the
transmitted vibration cancellation signals.
SUMMARY OF THE INVENTION
[0024] It is an object of the present invention to provide a
controller for an active control system for simultaneously reducing
both vibration and noise in aircraft applications, and particularly
helicopters.
[0025] Another object of the present invention is to provide a
controller for an active device and system for isolating the main
rotor system of a helicopter from the airframe for preventing the
low frequency vibration generated by the main rotor system from
being transmitted to the airframe.
[0026] A further object of the present invention is to provide a
controller for an active vibration reduction system for passing the
primary flight loads of the helicopter from the main rotor system
to the airframe while maintaining the average static position of
the gearbox relative to the airframe.
[0027] According to the present invention, a control system is
provided for reducing vibration generated by a vibrating plant, the
vibrating plant including a vibrating component, a structure and a
mount for mounting the vibrating component to the structure. The
control system comprises means for producing controlled vibrations
within the mount. Sensors are provided for sensing the current
position of the controlled vibration producing means, the vibration
being transmitted through the mount from the vibrating component to
the structure, and at least one of the characteristic frequencies
at which the vibrating plant operates and developing signals
indicative thereof. A first controller receives as input the signal
from the position sensor located on the controlled vibration
producing means and generates an output signal. A second controller
receives as input the transmitted vibration sensor signal and the
plant rotational sensor signal and generates an output signal.
Means are provided for combining the output signals from the first
and second controllers into a control signal for controlling the
vibration producing means such that the vibration transmitted from
the vibrating component to the structure through the mount is
reduced.
[0028] Further according to the present invention, a control system
is provided for an active system for reducing vibration generated
by a vibrating plant, the vibrating plant including a vibrating
component, a structure and a hydraulic mount for mounting the
vibrating component to the structure. The control system comprises
at least one hydraulic actuator for producing controlled vibrations
within the mount. Sensors are provided for sensing the current
position the hydraulic actuator relative to the mount, vibrations
being transmitted from the vibrating component through the
hydraulic mount to the structure, and at least one of the
characteristic frequencies at which said vibrating plant operates.
The sensors produce signals representative thereof. A fixed, low
bandwidth, near-DC, proportional/integral/derivative (PID)-based
broadband control compensation feedback position controller
utilizes the position sensor signal to produce position control
signals to minimize the offset between the sensed hydraulic
actuator position and a predetermined hydraulic actuator position.
An adaptive Filtered-X least-mean-square (LMS) based narrow-band
vibration controller utilizes the vibration sensor signal to
produce vibration control signals at multiple frequencies of the
sensed plant characteristic frequencies. Means are provided for
combining the position control signals with the vibration control
signals and generating an output signal which the hydraulic
actuator responds to for producing controlled vibrations in the
mount for reducing vibrations transmitted through the mount from
the vibrating component to the structure.
[0029] Also according to the present invention, a control system is
provided for active vibration reduction in a rotary wing aircraft
including an airframe and a main rotor system having an engine, a
rotor and a transmission gearbox mounted to the cabin support beam
located at the top of the airframe by at least one hydraulic mount.
The gearbox converts the engine force into the rotational force of
a rotorshaft. The control system comprises at least one hydraulic
actuator for producing controlled vibrations within the mount.
Sensors are provided for sensing the current position of the
hydraulic actuator, vibrations being transmitted from the main
rotor system through the mount to the airframe, and for sensing the
rotational frequency of the rotorshaft and producing signals
representative thereof. A fixed, low bandwidth, near-DC, broadband
control compensation feedback position controller utilizes the
actuator position sensor signal to produce quasi-static position
control signals to minimize the offset between the sensed hydraulic
actuator position and a predetermined hydraulic actuator position.
An adaptive Filtered-X LMS based narrow-band vibration controller
utilizes the vibration sensor signal to produce vibration control
signals. Means are provided for combining the position control
signals with the vibration control signals and generating an output
signal which the hydraulic actuator responds for producing
controlled vibrations in the mount for reducing vibrations
transmitted through the mount from the main rotor system to the
airframe.
[0030] A feature of the actuator position controller is the
attenuation output signals are maintained within a maximum range to
which the vibration producing means is capable of responding. The
extent of actuation of the actuator is thus maintained around a
predetermined point, preferably a center point, to ensure the
actuator has sufficient authority to respond to the vibration
cancellation signals. The position controller includes a scaling
function, a band elimination function, an objective function and a
compensation function. In one embodiment, the compensation function
produces the position control signal utilizing proportional,
integral, derivative control compensation.
[0031] The vibration controller features a frequency filter
adaptive to isolate sensed vibration signals at frequencies which
are multiples of the sensed characteristic or rotorshaft rotation
frequency, an objective function characterizing the magnitude of
the isolated signals, a compensation function producing a
correlation between the isolated signals and the control signal for
the controlled vibration producing means, and an adaptive filter
which generates attenuation output signals minimizing the isolated,
correlated signals. In one embodiment, the frequency filter
comprises a band-pass filter and a notch filter receiving as input
the characteristic or rotorshaft rotation frequency, the notch
filter adapting its filter window based on the input frequency. The
vibration controller also features a frequency downshift function
which converts the vibration sensor signals to signals at baseband
DC and a frequency upshift function which converts the baseband DC
signals into in-band, attenuation path-based, vibration control
signals. The vibration controller may also include an input
function which performs antialiasing and scaling functions on the
vibration sensor signals, a normalization function which normalizes
the isolated signals, an output function which scales the vibration
control signals, and a weight limiting function which evaluates the
vibration control signals and transmits a freeze signal to the
adaptive filter function affecting the adaptive abilities of the
adaptive filter function.
BRIEF DESCRIPTION OF THE DRAWINGS
[0032] For a more complete understanding of the present invention,
reference should now be had to the embodiment(s) shown in the
accompanying drawing(s) and described below. In the drawings:
[0033] FIG. 1 is a schematic representation of a helicopter
transmission arrangement;
[0034] FIG. 2 is a schematic representation of an embodiment of a
system for reducing vibration and noise passing from a helicopter
transmission gearbox to the airframe;
[0035] FIG. 3 is a schematic representation of a control system
according to the present invention shown in an embodiment of an
active system for vibration reduction with two control paths;
[0036] FIG. 4 is a schematic representation of an embodiment of a
saturation prevention position control path in a control system
according to the present invention;
[0037] FIG. 5 is a schematic representation of an embodiment of a
vibration attenuation control path in a control system according to
the present invention; and
[0038] FIGS. 6A and 6B are a flow diagram of an embodiment of a
control system according to the present invention with one
vibration attenuation control path and one saturation prevention
position control path.
DESCRIPTION
[0039] Certain terminology is used herein for convenience only and
is not to be taken as a limitation on the invention. For example,
words such as "upper", "lower", "left", "right", "horizontal",
"vertical", "upward", "downward", "clockwise" and
"counter-clockwise" merely describe the configuration shown in the
FIGs. It is understood that the components may be oriented in any
direction in the terminology. Therefore, the present invention
should be understood as encompassing such variations unless
specified otherwise.
Overview of the Invention
[0040] The "Controller" according to one embodiment of the present
invention is a signal processing-based and software-based
electronics system that receives vibration sensed from a vibrating
plant, including a vibrating component and structure, and controls
at least one device for inducing changes within the plant to reduce
the sensed vibrations. The Controller reduces vibration by
combining control signals generated by multiple processing paths
into output control signals. At least one of the processing paths,
the "Attenuation" path, generates a dynamic control signal
utilizing a narrow-band feedback control algorithm. At least one
other of the processing paths, the "Saturation Prevention Position
Control" path, generates a quasi-static control signal implemented
through a fixed, low bandwidth, broadband feedback algorithm. The
Attenuation processing path generates signals which drive the
controlled devices to reduce the sensed vibrations. The Saturation
Prevention Position Control path generates signals intended to
prevent the controlled devices from reaching saturation positions
such that the controlled devices would no longer be able to respond
effectively to the Attenuation path signals. In addition, the
Saturation Prevention Position Control path maintains the relative
position between the vibrating component of the plant and a desired
non-vibrating portion of the plant. In certain embodiments of the
present invention, only vibrations at certain frequencies are
desired to be controlled. Those frequencies are referred to as the
"frequencies of interest".
[0041] According to the present invention, functional operations
are performed by software executing on a digital signal processor
(DSP). Such DSP's are commercially available and include the TMS
320C30 floating point processor. The DSP's typically include a
central processing unit (CPU) for execution instructions and
performing arithmetical operations, random access memory (RAM) for
storing instructions and program data, programmable read only
memory (PROM) for storing static data such as program instructions,
clock circuitry, and mass storage devices such as disk drives or
tape drives. In the following description, reference is made to
data being transferred between the functional components of the
Controller. Such transfer may involve the sending of signals
between electrical components. Such transfer may also include not
only the moving of data within the RAM, but also any other method
by which one function can indicate to another function the location
of data. Such transfer methods include providing the address of
data within the RAM. Reference is also made to connections between
functional operations of the Controller. Such connections need not
necessarily be physical wiring connections, but can represent the
flow of control as the DSP executes the program instructions.
[0042] The output control signals generated by the Controller drive
an actuation system including one or more controlled devices. These
controlled devices are attached to the vibrating component of the
plant and respond to input signals by producing some change upon
the plant which alters the vibrational characteristics of the
plant. One example of a controlled device is an active spring that
sits on the mounting location between a vibration generating plant
and some attached body. Such a spring would respond to signals
generated by the Controller by either stiffening or adjusting the
resonant frequencies at which the spring would vibrate which would
alter the vibrations being transferred through the spring to the
attached support structure. Another example of a controlled device
is an active transmission mount located in series between the
transmission of a rotary wing aircraft and the airframe. The active
transmission mount includes hydraulic actuators which respond to
the Controller by altering the degree of hydraulic actuation of the
transmission relative to the airframe, thus altering the vibrations
transmitted through the active transmission mount. Other means for
altering the vibration characteristics of a plant are well known in
the art.
[0043] Referring now to the drawings, wherein like reference
numerals illustrate corresponding or similar elements throughout
the several views, FIG. 1 illustrates a transmission arrangement 20
for a helicopter. The transmission arrangement 20 includes a
gearbox 22 which is connected to a helicopter rotor head (not
shown). The gearbox 22 is also connected to the drive train 24 of
the helicopter's engine 26. The gearbox 22 is supported by an
airframe comprising a structural element 28. The gearbox 22
includes a plurality of mounting feet 30 which are attached to the
airframe structure 28. Active transmission mounts (ATMs) 32 are
mounted in series between each gearbox mounting foot 30 and the
airframe structure 28 for isolating the mounting feet 30 of the
main rotor gearbox 22 from the airframe.
[0044] The ATM 32 is a part of an active transmission mount system
34, an embodiment of which is schematically illustrated in FIG. 2.
In FIG. 2 the ATM system is viewed looking down at gearbox 22. The
ATM system 34 comprises one or more hydraulic ATM actuators 36
associated with each of four ATMs 32, a plurality of sensors 38-42
positioned throughout the aircraft, and an electronic Controller
100 which sends signals to a hydraulic actuation system 46 for
commanding the actuation system to actuate the ATM actuators 36
according to the desired operational state. For simplicity only a
single ATM actuator is shown connected to each ATM 32 and
transmission foot 30.
[0045] The sensors comprise position sensors 38 for monitoring the
static position of the feet 30 relative to the airframe 28. These
sensors 38, shown for a single actuator 36 and gearbox foot 30
combination in FIG. 2, are used along with the Controller 100 and
the actuation system 46 to ensure that the transmission does not
move out of static alignment with other elements of the airframe.
The preferred location and type of sensors 40, 42 are a function of
the type of control approach used by the Controller 100. For
example, one type of control approach utilizes sensors 40 that are
located adjacent to the mounting feet 30 and the ATM's 32. These
sensors 40 comprise accelerometers to sense airframe acceleration.
This same control approach may use pressure sensors 42 to sense
dynamic pressure fluctuations in the actuator fluid lines 50. An
alternate control approach may use accelerometers 40 mounted at
selected locations within the airframe, such as at the foot of the
pilot or a seat. The choice of local sensors (accelerometers 40 or
pressure sensors 42) or remote accelerometers 40 is largely based
on the type of airframe to which the ATM system 34 is applied and
is also based on the stiffness requirements defined for the ATM
actuators 36. Other sensors 39, 41 are located on the main rotor
shaft 43 for measuring the rotational rate of the shaft 43.
[0046] The signals output from the sensors 38-42 are provided for
processing to the Controller,100. For each mounting foot 30, the
Controller 100 determines the position of the foot 30 and vibratory
loading of the airframe based on the sensed signals being
transmitted by the sensors. The Controller 100 then determines a
desired operational state for each ATM actuator 36 as a function of
one or more of the sensed signals and operates to nullify position
offset of the gearbox 22 while also reducing the vibratory load
passing through the ATMs 32 and into the airframe.
[0047] The hydraulic actuation system 46 supplies a hydraulic fluid
under pressure to each hydraulic actuator 36 so that the actuator
moves in the desired manner and at the desired frequency to reduce
the sensed vibrations emanating from a mounting foot 30 of the
gearbox 22 passing into the airframe. In the illustrated
embodiment, the hydraulic actuation system 46 includes one or more
electro-hydraulic valves 48 which are each electrically connected
to the Controller 100 via a control line 52 for supplying current
to the valve 48. For example, the hydraulic inputs of two of the
actuators 36 shown in FIG. 2 are interconnected into a common
hydraulic fluid line 50 and connected to the hydraulic control
valve 48. For the sake of simplicity, only a single hydraulic
control valve 48 and associated hydraulic interconnections are
shown. The Controller 100 generates output control signals to the
hydraulic control valve(s) 48 in response to the signals received
from the sensors 38-42. The valve(s) 48 open and close in response
to the output control signals to provide a vibratory flow of high
pressure hydraulic fluid from a fluid source (not shown) to and
from the actuators 36. In FIG. 2, the supply flow into the valve 48
is generally indicated at 54. The hydraulic pressure and location
of the actuators' pistons are thus adjusted by the Controller 100
based on the signals from the sensors 38-42.
[0048] The active transmission mount system 34 of the present
invention acts to isolate the vibratory and acoustic loads
generated by the main rotor gearbox 22 from the airframe. The ATM
system 34 achieves vibration reduction by controlling the applied
fluid flow within the ATM actuators 36, and thus the hydraulic
pressure acting on the pistons in the actuators 36. A quasi-steady
pressure is applied to each actuator 36 to react to the applied
quasi-steady flight and maneuvering loads. The vibratory loads that
are applied along the actuator's principle, or "active", axis are
transmitted into the hydraulic column. This causes cancellation of
pressure fluctuations which would otherwise be transmitted into the
airframe causing vibration if left unaltered. Generally, an
increase in hydraulic pressure on the pistons when a vibratory load
pushes on the actuator 36 is relieved by the ATM system 34 by
removing fluid, and a decrease in hydraulic pressure when a
vibratory load pulls on the actuator 36 is accommodated by the ATM
system 34 by increasing hydraulic fluid flow to the actuator.
Hence, the actuator 36 is operated by removing and supplying a
sufficient amount of hydraulic fluid against the head of the piston
to allow the piston to translate in substantially the same
direction and at substantially the same frequency as the vibrating
gearbox 22. In this way, the ATM system 34 allows relative motion
between the gearbox 22 and the airframe at low vibration
frequencies, typically greater than about 2 Hz, so that the gearbox
22, in effect, floats in a dynamic sense with respect to the
airframe, but maintains a steady, static position relative to the
airframe. As a result, vibratory pressure is minimized, thereby
reducing the transfer of vibration related to the applied rotor
vibratory loads from the ATM 32 to the airframe.
[0049] The Controller 100 functions within the ATM system to reduce
vibrations by combining two control signals which are generated by
a vibration controller and a position controller, respectively, for
each hydraulic control valve 48. The vibration controller provides
a dynamic control signal utilizing a narrow-band feedback control
algorithm. The position controller provides a quasi-static control
signal implemented through a fixed, low bandwidth, broadband
feedback algorithm. The position controller seeks to keep the
position of the actuators 36 at a desired point located near the
midpoint of the actuator's displacement range. Keeping the
transmission foot from reaching the "hard-stop" position on the
actuators 36 allows optimal dynamic narrow-band isolation of low
frequency vibrations (20 Hz to 60 Hz).
[0050] Actuator "hard-stops" define the minimum and maximum extent
of the actuator's displacement range. Keeping the transmission foot
from reaching the "hard-stop" position on the actuators 36 also
allows optimal passive, narrow band isolation of high frequency
noise (e.g., >500 Hz) with a passive noise isolator 58. The
passive isolator 58 introduces softness into the hydraulic system
at predetermined frequencies to allow the system to attenuate high
frequency and low amplitude, {fraction (1/1000)} inch, noise that
is otherwise transmitted by the gearbox feet 30 to the ATM 32
causing high frequency noise in the fluid lines 50 which, in turn,
leads to noise in the aircraft. In order to reduce this high
frequency noise, the hydraulic line 50 is connected to the passive
isolator 58 which in one embodiment is one or more tuned stubs. It
is understood that the other hydraulic lines (not shown) that
interconnect the valves 48 and their associated actuators 36 are
also connected to passive isolators.
[0051] Referring now to FIG. 3, the Controller 100 for the ATM
comprises an ATM vibration controller (ATMVC) 200, an ATM position
controller (ATMPC) 400, an ATMVC pre-run 210 and an ATMPC pre-run
410. The ATMVC 200 is responsible for providing vibration control
output signals on line 206 to cancel sensed vibrations. Thus, the
ATMVC 200 serves to attenuate vibrations. The ATMVC 200 vibration
control output signals are based upon vibration sensor input
signals received from the accelerometers 40 and the rotational rate
input signal received from the rotation sensors 39, 41 on the rotor
shaft 43.
[0052] The ATMPC 400 is responsible for providing position control
output signals on line 228 to keep the actuators 36 properly
centered around a desired position. Thus, the ATMPC 400 serves to
prevent the actuators 36 from "saturation", which occurs if the
extent of actuation of the actuators 36 approaches either the
minimum actuator extent or the maximum actuator extent. The
position control output signals are derived from the position
sensors 38.
[0053] The position controller output signal on line 228 and the
vibration controller output signal on line 206 are combined using
digital summing 62. The resulting Controller 100 output on control
line 52 is converted from digital to analog form and serves as the
input signals to the hydraulic control valves 48 connected
hydraulically to one or more actuators 36.
[0054] The ATMPC Pre-Run 410 and ATMVC Pre-Run 210 are active only
during the initial set-up training phase of the Controller 100.
During this set-up training phase, sample control outputs are sent
to the hydraulic control valves 48. The Controller 100 correlates
changes in sensed vibrations from the sensors 40 with each sample
output. The Controller uses this correlation to build a model of
the signal transfer characteristics between the actuator 36 and
sensor 40. This signal transfer characteristic is called the
actuator-to-accelerometer transfer function. This model is then
used to help calculate the vibration reduction control signals
generated by the Controller 100 in its normal operational mode
which are transmitted to the hydraulic control valves 48 on control
line 52. Preferably, the model of the actuator-to-acceleromete- r
transfer function is created by using the ATMVC Pre-Run 210 as a
structural-probing function and calculating a model based on
measurements from the probing signals. It is understood that
alternate methods of creating models would suffice for the purpose
of obtaining an actuator-to-accelerometer transfer function. These
alternate methods include building equations into the Controller
100 based on mathematical or theoretical models on how a plant
behaves.
[0055] The ATMPC Pre-Run 410 determines the signal voltage levels
required to drive the actuator 36 to each opposing hard-stop. From
these voltage levels, a centering voltage is calculated. For the
case shown in FIG. 2 with two actuators 36 driven from a single
servo valve 48, the centering voltage calculation is based on the
minimum voltages required to drive the two actuators 36 to their
respective hard-stops. The ATMPC Pre-run-410 and the ATMVC Pre-Run
210 are run sequentially.
[0056] FIGS. 4, 5 and 6 depict an embodiment of a Controller 100
according to the present invention. It is understood that the
present invention could be implemented using other algorithms for
the attenuation and saturation prevention position control paths.
Therefore, nothing in the description of the following embodiment
is intended to limit the present invention to only those described
for this specific implementation. One skilled in the art could
easily implement alternatives to these specific algorithms that
have the same effect of controlling vibration while preventing the
controlled devices from reaching control extents which no longer
allow for full ranges of vibration control.
[0057] FIG. 4 illustrates a preferred embodiment of the ATMPC 400.
The position sensor 38 signal for each of the actuators 36 at the
input of the ATMPC 400 is first processed by a scaling and
pre-processing function 420 which applies signal level scaling and
analog filtering. The resultant signal is then processed through a
band-elimination function 440, removing those frequencies that are
not required for use for by the ATMPC 400. The resulting signal is
then compared to the signal level of a desired actuator position
455 in the summing box 450. The difference between these two
signals is sent to the position control compensator function 460
which provides spectrum equalization. This signal is sent to an
output scaling function 480, providing output gain control.
Preferably, the ATMPC 400 uses a fixed, low bandwidth, near-DC,
broadband control compensation feedback algorithm known as the
proportional/integral/derivative (PID) controller approach. The
ATMPC 400 provides a quasi-static position controller output signal
on line 228.
[0058] FIG. 5 illustrates a preferred embodiment of the ATMVC 200.
The vibration sensor 40 signals at the input of the ATMVC 200 are
first processed by a scaling and pre-processing function 220 which
applies signal level scaling and analog filtering. This is the same
signal level scaling and analog filtering used in the ATMPC 400.
The resultant signal is then processed through a spectrum band-pass
function 240, allowing only the spectrum of frequencies necessary
for vibration control. The resultant signals are normalized in a
normalizing function 260, and frequency downshifted in a down
shifting function 270 which feeds into an objective function 300.
The signals are then processed in a compensation and adaptation
function 320 including an adaptive filter which seeks to minimize
the vibrations being sensed and processed. The adaptive filter
outputs the signals to an upshifting function 360 and then to an
output scaling function 370, providing spectrum equalization and
output gain control, respectively. A preferred adaptive filter for
use in the compensation and adaptation function 320 uses a
Filtered-X LMS algorithm approach together with a narrow-band
compensation approach to achieve narrow-band feedback control. This
approach is robust since it allows the controller to adapt to
changes in the vibration control actuator-to-accelerometer transfer
function. The ATMVC 200 path provides a dynamic vibration
controller output signal on line 206.
Position and Vibration Control Common Processing Path
[0059] Referring now to FIGS. 6A and 6B, a detailed description of
a preferred embodiment of the Controller 100 according to the
present invention is presented. FIGS. 6A and 6B show control
processing for a 2-dimensional system for the control of two
hydraulic control valves 48. A complete ATM system 34 for a rotary
wing aircraft with four transmission feet 30 may have a dedicated
hydraulic control valve 48 for each of eight or more ATM actuators
36. The input position sensor 38 (FIG. 6A) and vibration sensor 40
signals are processed initially in a common processing path. This
common processing path contains the ATMPC scaling and
pre-processing function 420 and ATMVC scaling and pre-processing
function 220 discussed above. Within this common path, the sensor
signals are processed by a data acquisition Unit (DAU) 160, gain
control function 148 and a sensor combining function 150. The DAU
provides anti-alias filtering 162, scaling and conversion from
analog to digital format 164. In a preferred embodiment, these
functions are performed by a single DAU 160.
[0060] For anti-aliasing purposes, the DAU 160 must filter out
(i.e., eliminate) frequencies higher than the Nyquist folding
frequency and deliver digital values at the execution rate of the
Controller 100. Nyquist frequency analysis is well known within the
art. Additionally, the DAU 160 should provide a sample with enough
accuracy, preferably 16-bit digital samples, for processing, along
with programmable analog gains and sufficient anti-aliasing filter
types. A suitable DAU 160 is a Tustin Series 2100 Data Acquisition
Unit System produced by Tustin Electronics Company of Anaheim,
Calif., U.S.A.
[0061] In an ATM system 34 for a rotary wing aircraft, the
Controller 100 preferably executes at 512 Hz. This sampling rate
allows for the potential of processing signals above the targeted
highest frequency of 12P or 60 Hz. Therefore, a filter 162
frequency of 200 Hz may be used, which is below the Nyquist folding
frequency of (512 Hz/2)=256 Hz. For anti-aliasing purposes, the
preferred DAU 160 incorporates a bank of 8-pole/8-zero elliptic low
pass filters with a cutoff frequency of 200 Hz. The cutoff
frequency is programmed into the unit via the digital configuration
table. The 8-pole/8-zero elliptical low-pass filter has desirable
qualities such as linear phase delay over the vibration control
bandwidth in the frequencies desired to be reduced as well as
providing sufficient attenuation in the stop band. Linear phase
delay over the bandwidth of interest translates in a simple, and
undistorted, time shift of the original vibration control signal,
which contains the frequencies of interest. Also, the magnitude of
this time delay is small enough not to interfere with the position
controller feedback loop. Regarding the stop band attenuation, the
8-pole/8-zero elliptical low-pass filter provides large attenuation
in order to minimize effects for signal aliasing. This filter type
is effective for small frequency ranges, particularly in the
vibration control frequency band between 20 Hz and 65 Hz. In
applications other than rotary wing aircraft in which the bandwidth
of the signals that the vibration control path seeks to reduce
differs considerably, anti-aliasing filters appropriate to the new
bandwidth would replace the existing elliptical filters.
[0062] The Tustin 2100 DAU has a programmable gain feature which
allows the user to specify a 16-bit word for each input channel
which controls the gain level for that channel. Each channel
corresponds to a single vibration or position sensor. The Tustin
2100 allows the user to use 12 bits as a gain amplitude factor. It
is an engineering principle that a 2.times. voltage increase=6 dB.
This principle is discussed in the book Introduction to Electrical
Engineering, Authors: J. David Irwin, David V. Kems, Jr.,
Publisher: Prentice-Hall (1995); ISBN: 0023599308. With 12 bits,
where each bit represents a 2.times. increase, the amplitude gain
can range from 0 to 66 dB.
[0063] The values used to program the Tustin DAU for anti-aliasing
and gain purposes are stored in a digital configuration table. The
data within the digital configuration tables includes the 200 Hz
filter cutoff frequency and a gain factor for each input channel.
Some sensors may have an input signal on the order of millivolts.
By using the analog gain feature provided by the Tustin DAU to
scale the sensor signals up to approximately the +/-5 volt range,
more of the full useful dynamic range of the A/D unit from 0 to
+/-10 volts are utilized when the signals are converted to digital
form. This scaling permits greater accuracy in reading sensor
signals to be processed later, but are not a necessary component of
the present invention. For sensors that supply input signals with
larger magnitudes, scaling is unnecessary.
[0064] Although the above description uses a single DAU 160 which
incorporates multiple functions, the anti-aliasing 162, scaling and
A/D conversion 164 need not be done within a single device.
Multiple devices could easily be arranged to perform the same
functions as the preferred DAU performs. The processed signals from
the DAU 160 are then sent to the change gain function 148.
[0065] The change gain function 148 equalizes the magnitude of the
vibration sensor signals 40 and the position sensor signals 38
separately. The change gain function 148 multiplies the input
digital samples in the discrete time-domain with a set of digital
coefficients stored in memory in, for example, a coefficient table.
Each sensor input is multiplied with a single corresponding
coefficient within the coefficient table. This set of numbers
within the coefficient table is usually greater than or equal to
unity. The change gain function 148 provides the user with the
ability to maintain, upgrade and adapt the Controller 100. Due to
these advantages, a preferred embodiment of the present invention
includes a change gain function 148.
[0066] The sensor combining function 150 separates the filtered and
digitized input sensor 38 and 40 signals into position and
vibration signals. Preferably, this separation is done implicitly
by arranging the vibration and position sensor signals in memory
and then passing the address of the vibration sensor signals to the
vibration processing path and the address of the position sensor
signals to the position processing path.
Vibration Control Processing Path
[0067] After the common vibration and position sensor signal
processing, the vibration signals are processed by the vibration
control processing path (VCPP). The VCPP can conceptually be broken
into three phases: the preprocessing phase consisting of the band
pass function 240 and the normalizing function 260; the vibration
control phase consisting of the downshifting function 270, the
objective function 300, the compensation and adaptation function
320 (FIG. 6B), and the upshifting function 360; and the
post-processing phase consisting of the output scaling function 370
and the weight limiting function 500.
[0068] Preprocessing the Vibration Input Sensor Signals
[0069] After the sensor combining function 150 (FIG. 6A) has
separated the vibration and position sensor signals, the vibration
signals are filtered by the band-pass function 240 to pass only
signals with frequencies which are desired to be controlled. In the
preferred embodiment, this filtering is accomplished with a
band-pass filter 242 followed by a notch 1P filter 244. In some
applications, the frequencies of interest are known before
operation and no measurement during operation is necessary for the
Controller 100. In other applications, the frequencies of interest
have a relationship to some base rate at which the associated
vibration-generating component operates. Measurement of this base
rate permits the Controller 100 to isolate and control vibrations
of the frequencies of interest.
[0070] In the rotary wing aircraft application, the base rate is
the rotorshaft rotational rate P, and the measurement is performed
by the rotation sensors 39, 41. The rotation sensors 39, 41 may
comprise one tachometer with two TTL-compatible analog inputs which
measure the rotorshaft rotational rate P. A 1-per-rev tachometer 39
permits a mean estimate of the IP frequency while a 1024-per-rev
tachometer 41 provides an enhanced means to estimate both the mean
and variance of the IP frequency. The Tustin DAU 160 provides two
separate TTL-compatible inputs, one for each of the tachometer
signals. A tachometer board 505 examines the TTL signals received
by the DAU 160 and provides to this input a number representing the
rotorshaft rotational rate, P. This number is represented by fo
adjacent signal line 516. The use of multiple inputs in a
tachometer provides redundancy in case one tachometer input
fails.
[0071] The band-pass filter 242 passes only those signals with
frequencies of interest and preferably comprises a parallel
connection of second order sections. The band-pass filter design is
done off-line using filter design techniques using as parameters a
512 Hz sampling frequency, cutoff frequencies and the number of
second order sections to use. The parameters used to design the
filter are dependent upon the required filter response and would be
evident to one skilled in the art.
[0072] After the band-pass filtering 242, the filtered sensor
signal is transmitted to the notch IP filter 244. The notch IP
filter 244 is useful after the band pass filtering step 242 because
the notch IP filter 244 provides a much higher attenuation of the
base rate (e.g., 1P) frequency than the band-pass filter 242. The
notch 1P filter removes the input signal corresponding to IP. In
the rotary wing aircraft application, the IP frequency needs to be
filtered to prevent any influence on the primary control frequency.
The design of the notch IP filter 242 assures no phase shift of
control frequencies of interest (i.e., 4P, 8P and 12P) which lie
outside the IP frequency of attenuation.
[0073] The input to the notch 1P filter 244 on signal line 516 is a
measurement of the base rate at which the controlled plant is
operating. In embodiment shown in FIG. 6, this measurement is
received from the tachometer board 505 described above. This
measurement updates the coefficients of the notch 1P filter 244 to
keep the filter focused around the current IP frequency as closely
as possible. In the embodiment shown in FIG. 6 the coefficients for
the notch filter 244 are not updated based upon the tachometer
reading because the plant's base rate, the rotorshaft rotational
rate, does not change appreciably during operation. It is
understood that for a rotary wing aircraft application with varying
base rates of a few percent, modification of the notch filter 244
coefficients based on the associated plant's base operational rate
is preferred and is within the present invention.
[0074] The band-pass filter 242 and notch IP filter 244 could be
modified or combined as necessary depending on the set of
frequencies of interest. Other filters could be implemented as long
as the result is a signal in which the frequencies of interest are
present and frequencies not desired are sufficiently attenuated
after filtering.
[0075] Band pass filter design is well known. One method of
creating the band pass filter 242 is by using a software tool
called Matlab produced by The MathWorks, Inc., of Natick, Mass.,
U.S.A. Wave digital filter techniques may be used to design notch
filters. Design methodologies for such filters are commonly known
in the art. References that detail such methodologies include:
[0076] 1. Circuits and Systems, A Modem Approach, written by
Athanasios Papoulis and published by Holt, Rinehart and Winston
(1980) (ISBN 0030560977).
[0077] 2. Digital Filters: Analysis and Design, written by Andreas
Antoniou and published by: McGraw Hill (1993) (ISBN
007002121X).
[0078] After filtering the vibration sensor signals to pass only
the frequencies of interest, the vibration sensor signals are
transmitted to the normalizing function 260. Preferably, the
normalizing function 260 comprises simple vector scaling. Each
input sensor signal value is multiplied by a predetermined scaling
factor. This scaling factor is determined to provide an equivalent
level of 1 Volt in its digital 16-bit representation implemented in
a 32-bit floating point numeric processing architecture. This level
is empirically determined to facilitate a smoother conversion of
the adaptive section.
[0079] The normalizing function 260 aids the user's ability to
maintain, upgrade and adapt the ATM Controller 100. The normalizing
function 260 facilitates these goals by making the system more
modular. Due to these advantages, the preferred embodiment of the
present invention includes the normalizing function 260.
Vibration Control Phase
[0080] The Vibration Control Phase includes the downshifting
function 270, the objective function 300, the compensation and
adaptive function 320 (FIG. 6B) including an adaptive Filter 325,
and the upshifting function 360. The Vibration Control Phase is
executed separately for each frequency of interest. In the rotary
wing application, the frequencies of interest are 4P, 8P and 12P.
For simplicity, FIGS. 6A and 6B show only the Vibration Control
Phase at 4P. Many methods of separate execution are well known in
the art and include multi-tasking, sequential execution and
parallel execution on separate processors. The present invention
does not depend on the method of separate execution other than the
requirement that the processing for all frequencies of interest
must be completed within the one execution cycle. In the embodiment
shown in FIGS. 6A and 6B, the Controller 100 executes at 512 Hz.
Preferably, the processing for the frequencies of interest is
executed on the same Digital Signal Processor (DSP) in a sequential
manner. In the following description, the term "NP" is used to
represent the individual frequency being processed for each of the
frequencies of interest.
[0081] Referring to FIG. 6A, the downshifting function 270 performs
a frequency downshift of the vibratory load signal at NP to
base-band. Base-band in the preferred embodiment is DC (frequency=0
Hz). An additional input to this function is the f.sub.0 parameter
from the tachometer process received on signal line 516, from which
NP is computed and used to generate a complex sinusoidal look-up
table which performs the NP frequency downshift. The downshift
corresponds to the following formula:
baseband=FILTERED SIGNAL*(e.sup.-j.omega.n)
[0082] where
[0083] FILTERED SIGNAL=vibration control signal filtered through
the frequency filters described above; and
[0084] "j` is the imaginary number 1 ( j = ( - 1 ) 2 )
[0085] and
[0086] (e.sup.-j.omega.n)=cos(.omega.n)-j sin(.omega.n) and
[0087] .omega.=the angular frequency currently being processed
(i.e. 4P/8P/12P in the preferred embodiment) in
radians/sec=2.times..pi..times.- NP
[0088] n=the current sample count.
[0089] The above equations require calculating sine and/or cosine
values. In the embodiment shown in FIGS. 6A and 6B, the main IP
frequency component is very steady with negligible variance. Thus,
it is possible to construct an "a priori" sine wave table in two
steps. The sine wave table supplies the necessary sine and cosine
values.
[0090] In the first table creation step, a table of length M
floating point numbers (32-bit) is created. In the embodiment shown
in FIGS. 6A and 6B, M=5120 elements. This length provides a
sinusoidal look-up capability with a resolution .function..sub.B of
512 Hz/5120 element=0.1 Hz per element. Different resolutions
(coarser or finer) can be achieved with a different table length
within the limits of the DSP memory resources. It is understood
that as long as the variance of the base frequency is small, a
coarse table resolution can be constructed. Otherwise, a larger
table needs to be constructed in order to keep .function..sub.B
small.
[0091] In the second table creation step, the table of length M is
populated one entry at a time using the following equation:
cos(2.pi.{.function..sub.B/.function..sub.S}Nn) (A)
[0092] where
[0093] f.sub.B=Resolution Frequency in Hz,
[0094] f.sub.S=Sampling frequency in Hz,
[0095] N=harmonic order (1,2, . . . ,int(f.sub.S/2)),
[0096] n=0, 1, 2, . . . ,(M-1).
[0097] Thus, for N=1, each sample of the table is used to produce
.function..sub.B, and in order to produce N*.function..sub.B, every
other N sample is used to construct it. For example, for
.function..sub.B=0.1 Hz, the 4P frequency component (20.8 Hz) would
be obtained using N=208. In this case, every other 208.sup.th
sample is used to generate the 4P signal.
[0098] The lookup into the sinusoidal table is done by
calculating
FrequencyJump=CurrentFrequency/Resolution (i.e.,
NP/.function..sub.B)
CurrentIndex=(CurrentIndex+FrequencyJump) MOD TableSize
[0099] Note that a unique CurrentIndex must be stored in memory for
each separate controlled frequency and its value must be retained
between execution cycles.
[0100] Another application of the present invention may allow for a
precise sinusoidal generator based upon the following second order
recursive equation: 2 x [ n ] = 2 cos ( 2 f f s ) x [ n - 1 ] - x [
n - 2 ] ( B )
[0101] where .function. is the oscillator frequency and
.function..sub.S is the sampling frequency which, in the embodiment
shown in FIGS. 6A and 6B, is 512 Hz. A reference for the
implementation of equation (B) is "Improving Performance of Digital
Sinusoidal Oscillators by Means of Error Feedback Circuits" IEEE
Transactions on CAS, Vol. CAS-33, No.4, April 1986, p. 373. Other
methods of sinusoidal generation are well known in the art and may
be used as desired based on processing, memory and resolution
constraints.
[0102] The downshifting function 270 creates a complex number with
both a real and an imaginary value for each of the sensor signals
of the form:
(FILTERED SIGNAL*cos(.omega.n)) and (FILTERED SIGNAL*(-j
sin(.omega.n))
[0103] This complex number is processed through the objective
function 300 and compensation and adaptation function 320 (FIG. 6B)
portions of the Vibration Control Phase. The real and imaginary
value are converted back to a single real number in the upshifting
function 360. The downshifting 270 and upshifting 360 functions
are, in effect, inverse operations in the time and frequency
domains, since the signal is multiplied by the complex exponential.
The functions perform inverse frequency shifting in the Fourier
domain. Note that the terminology of modulation and demodulation
are also used in the art for the downshifting and upshifting
functions described here.
Objective Function
[0104] The objective function 300 (FIG. 6A) permits the user to
emphasize reduction of the sensed values of certain sensors deemed
to be more effective in reducing target vibrations while
de-emphasizing those sensors deemed to be less effective in
vibration control performance.
[0105] In the embodiment shown in FIGS. 6A and 6B, the objective
function 300, "Q", comprises a default unity matrix (i.e., each
sensor 40 is equally weighted) or a variant thereof with selected
null entries for those sensors 40 which empirically do not
contribute to performance. The matrix "Q" is a square matrix with
the number of row/columns equal to the number of input vibration
sensor readings emanating from the downshifting function 270. The
matrix entries, or weights, are complex numbers determined based on
empirical analysis of the effect that each sensor's readings have
on overall vibration control performance. Within the objective
function 300 matrix, several of the imaginary matrix components are
0, but this is dependent on an analysis of the results of the
objective function 300.
[0106] The weights are selected to optimize performance based upon
input signal constraints, which are related to the squares of the
Q-matrix entries. The type of constraints that are defined in the
objective function 300, through the use of the Q-matrix, depend on
the inclusion of input vibration sensor 40 weightings according to
the following equation:
I.sub.j=.alpha..sub.j1V.sub.1[n]+.alpha..sub.j2V.sub.2[n]+. . .
+.alpha..sub.jKV.sub.K[n]j=1,2, . . . ,K (*)
[0107] Equation (*) above produces a particular constraint signal
I.sub.j through the selection of the .alpha..sub.ji. The V.sub.j[n]
represent the resultant signals from the downshifting function 270.
The equivalent matrix representation for equation (*), which
includes all constraint signals I.sub.j, is expressed by (**)
below: 3 I _ = _ V _ ; I _ = [ I 1 I 2 . . I K ] ; _ = ( 11 12 . .
1 K 21 22 1 K K1 K2 11 ) ; V _ = [ V 1 [ n ] V 2 [ n ] . . V K [ n
] ]
[0108] The coefficients .alpha..sub.ij are such that are
row-normalized as defined in the following equation: 4 _ j = 0 <
1 12 a j1 2 + a j2 2 + + a jK 2 1 (* **)
[0109] This expression (***) geometrically means the projection of
the vector {overscore (V)} on the vector .alpha..sub.j, which
expresses the optimization of the vector {overscore (V)} along the
objective vector .alpha..sub.j.
[0110] In the embodiment of the Controller 100 shown in FIGS. 6A
and 6B, these weights do not vary dynamically. Preferably, the
objective function 300 stores a separate matrix for each of the
frequencies of interest.
Compensation and Adaptation Function
[0111] Referring to FIG. 6B, the compensation and adaptation
function 320 consists of the compensation function 321 and the
adaptive filter function 325. The compensation function 321 allows
for the compensation of the actuator-to-sensor transfer function
measured in ATMVC pre-run 210. The compensation function 321 maps
the relationship between a given output control signal sent to the
associated servo valve 48 on signal line 52 and the effect on the
vibrations generated by the associated actuator 36 as measured by
each vibration sensor 40. The purpose of the compensation function
321 is to negate the phase shift and amplitude change resulting
from the response of the accelerometers 40 to controlled actuator
drive signals. Preferably, this correlation consists of a matrix of
complex numbers, the plant compensation matrix.
[0112] Plant Compensation Training
[0113] The compensation function 321 is determined during the
vibration prerun 210 execution. During vibration prerun 210, a set
of control signal outputs, called probe signals,) are generated and
sent to the hydraulic control valves 48 on control line 52. The
vibration prerun 210 then records inputs from the vibration sensors
40 which have gone through the vibration pre-processing phase
described above. These sensor 40 inputs are correlated with the
control signal outputs which were applied to the plant on control
line52.
[0114] The correlation is calculated using the Fourier sine-cosine
coefficients at the frequencies of interest which, in a rotary wing
aircraft application, are 4P, 8P, and 12P. The vibration prerun 210
generates a cosine series which is output to the controlled device,
that is, the hydraulic actuator 36 (not shown). A sine series is
also generated but not used as output. After a waiting period which
allows the plant to stabilize, the vibration prerun 210 accumulates
five running sums for each sensor 40:
[0115] Sum 1--The product of the cosine series and the preprocessed
sensor reading input (device response)
[0116] Sum 2--The product of the sine series and the preprocessed
sensor reading input
[0117] Sum 3--The square of the cosine series
[0118] Sum 4--The square of the sine series.
[0119] Sum 5--The product of the sine and cosine series
[0120] The in-phase and lagging terms are computed using these
sums. The procedure is to drive the first controlled device by the
generated cosine series and obtain the response of all sensors.
Each sensor's response constitutes one row within column one of a
response matrix. The second controlled device is then driven by the
generated cosine series and the responses recorded constitutes
column two of the response matrix. This process is continued until
all the controlled devices have been driven by the generated cosine
series, and thus, all the columns within a plant compensation
matrix are created. The compensation matrix is calculated as the
pseudo-inverse of the plant compensation matrix unless the number
of valves 48 equals the number of sensors 40, in which case a
square matrix inverse is used.
[0121] The user supplies the following parameters to the vibration
prerun 210 function:
[0122] f.sub.0--frequency (Hz)
[0123] A--signal amplitude to produce desired output (volts)
[0124] T.sub.r--ramp-up time to reach signal amplitude A
(seconds)
[0125] T.sub.w--wait interval before beginning calculations
(seconds)
[0126] T.sub.a--accumulation time (seconds)
[0127] n.sub.v--number of control output signals which equals the
number of controlled devices
[0128] n.sub.e--number of error sensors
[0129] In the embodiment shown in FIGS. 6A and 6B, the frequencies
are 4P, 8P and 12P and typical ramp-up time and wait intervals are
two seconds and accumulation time is 4 seconds. Based on these
parameters, for each frequency of interest, the vibration prerun
210 calculates:
[0130] Angular frequency
w.sub.0=2.pi..phi..sub.0
[0131] Total number of samples
n.sub.s=(T.sub.a+T.sub.w)f.sub.s (f.sub.s=sampling frequency)
[0132] No. of samples for ramp-up
n.sub.r=T.sub.rf.sub.s (f.sub.s=sampling frequency)
[0133] No. of samples to skip
n.sub.w=T.sub.wf.sub.s ramp-up) (f.sub.s=sampling frequency)
[0134] The accumulation values, yx.sub.1, yx.sub.2, xs.sub.1,
xs.sub.2, xs.sub.3, which correspond to Sum1, Sum2, Sum3, Sum4, and
Sum5, respectively are set to zero.
[0135] For each sample point on which a correlation will be made,
vibration prerun 210 will generate and output to the controlled
device the cosine series value:
x.sub.1(k)=A cos(kw.sub.0T.sub.s) k=0, . . . ,n.sub.s
[0136] generate the sine series value:
x.sub.2(k)=A sin(kw.sub.0T.sub.s) k=0, . . . ,n.sub.s
[0137] obtain preprocessed sensor input
y(k) for each sensor
[0138] Beginning at sample n.sub.w (end of wait period) vibration
prerun 210 will calculate and accumulate until n.sub.s (total
number of samples) the following sums described above:
[0139] Sum1:
yx.sub.1(k)=yx.sub.1(k-1)+y(k)x.sub.1(k)k=n.sub.w, n.sub.w+1, . . .
,ns
[0140] Sum2:
yx.sub.2(k)=yx.sub.2(k-1)+y(k)x.sub.2(k)k=n.sub.w, n.sub.w+1, . . .
,ns
[0141] Sum3:
xs.sub.1(k)=xs.sub.1(k-1)+x.sub.1(k)x.sub.1(k)k=n.sub.w, n.sub.w+1,
. . . ,ns
[0142] Sum4:
xs.sub.2(k)=xs.sub.2(k-1)+x.sub.2(k)x.sub.2(k)k=n.sub.w, n.sub.w+1,
. . . ,ns
[0143] Sum5:
xs.sub.3(k)=xs.sub.3(k-1)+x.sub.1(k)x.sub.2(k)k=n.sub.w,
n.sub.w+1,. . .,ns
[0144] After final sample n.sub.s, vibration prerun 210 will
calculate the following values:
[0145] Determinant
d=xs.sub.1(n.sub.s)xs.sub.2(n.sub.s)-xs.sub.3(n.sub.s).sup.2
[0146] In-phase term
a=(xs.sub.2(n.sub.s)yx.sub.1(n.sub.s)-xs.sub.3(n.sub.s)yx.sub.2(n.sub.s))/-
d
[0147] Out-of-phase term
b=(-xs.sub.3(n.sub.s)yx.sub.1(n.sub.s)+xs.sub.1(n.sub.s)yx.sub.2(n.sub.s))-
/d
[0148] Phase of response
.phi.=tan.sup.-1(b/a)
[0149] Magnitude of response
h.sub.0=sqrt(a.sup.2+b.sup.2)
[0150] Complex response
h=h.sub.0(cos .phi.-i sin .phi.); i=sqrt(-1)
[0151] The value h constitutes the row entry for the current sensor
input being processed in the column which represents the controlled
output device currently being driven by the generated cosine
series.
[0152] Once all sensors 40 have been processed for each controlled
device, a response matrix H has been calculated. The plant
compensation matrix is the inverse of H. If H is a non-square
matrix--the number of controlled devices does not equal the number
of sensors-then a pseudo-inverse matrix is calculated using the H
hermitian which is equal to the transpose conjugate of H.
Techniques for calculating the transpose conjugate are well known
in the art. The following set of equations defines how the plant
compensation matrix is calculated from the response matrix H in the
situation where there are two controlled outputs.
[0153] Matrix response
H=[h.sub.ij]: input i, output j
[0154] Compensation
If n.sub.v=n.sub.e: H.sup.-1 (matrix inverse of H)
[0155] Otherwise: H.sup.+ (pseudoinverse)
[0156] Condition No. of H.sup.HH See formula 2 for controlled
output case below
[0157] Pseudoinverse:
[0158] h.sub.ij=transfer function between error sensor i and valve
j.
[0159] H=response matrix with each element h.sub.ij containing the
transfer function between error sensor i and controlled output
j.
[0160] H.sup.+=(H.sup.HH).sup.-1H.sup.H (H.sup.H=H hermitian=H
transpose conjugate)
[0161] For a Controller 100 with more than two control output
control lines 52, the plant compensation matrix would be calculated
in a similar manner. In essence, this process is continued until
all of the controlled devices have been driven by the generated
cosine series, and thus, all the columns within the plant
compensation matrix are created, which, in general, is
non-square.
[0162] In the embodiment shown in FIGS. 6A and 6B, the vibration
prerun 210 function collects and stores the
actuator-to-accelerometer transfer function measurements for the
vibration control path, providing complex number entries for the
plant compensation matrix. The Controller 100 software allows the
user the option to store the individual sensor streams to an
external storage device such as a disk 215 (FIG. 6A). If the
individual sensor streams are stored to an external storage device
such as a disk 215, the plant compensation matrix could be
calculated by an offline function external to the Controller 100
which performs any equivalents to the above calculations. For a
rotary wing aircraft application, the plant compensation training
process is preferably performed periodically, every few seconds, to
account for changes in the actuator-to-accelerometer transfer
functions as the aircraft changes flight conditions.
[0163] Further, according to the embodiment shown in FIGS. 6A and
6B wherein the controlled device is a hydraulic actuator 36, the
level of the cosine series output to the actuator 36 during
vibration prerun 210 is manually adjusted to prevent the attached
plant from railing at the two piston hard-stops. Once an initial
adjustment is made, further manual adjustments before each
execution are unnecessary. This manual adjustment capability can be
modified to become automated for the preferred rotary wing
application.
[0164] Plant Compensation Processing During Vibration Control
Mode
[0165] The objective function 300 sends to the compensation
function 321 (FIG. 6B) function an array of weighted values. The
number of elements within the array is equal to the number of
vibration sensors. In a preferred embodiment, the compensation
function 321 multiplies the input array with the plant compensation
matrix resulting in an array with a number of elements equal to the
number of hydraulic control valves 48. The resultant array
represents a column vector of frequency downshifted, base-band,
compensated signals in complex number notation.
[0166] Adaptive Filter
[0167] The compensation function 321 outputs the resultant array
described above to the adaptive filter 325. The adaptive filter 325
seeks to generate output signals sent to the upshifting function
360 which will essentially minimize the values of the vibration
sensor signals 40 which have a base-band, compensated version of
the signal.
[0168] Preferably, the adaptive filter 325 is a Filtered-X Least
Mean Squares (Fx-LMS) filter. Feed-forward algorithms such as the
Filtered-X Least Mean Squares (LMS) algorithm minimize the measured
disturbance signals using a gradient descent algorithm to adapt the
coefficient of a FIR (Finite Impulse Response) filter. With
Feed-forward systems, the FIR filter coefficients are updated so
that the transfer function from the disturbance source to the
disturbance signals where cancellation is desired, is equal to the
net transfer function from the source through the reference sensor,
FIR filter, and actuator to the same disturbance signals. The
adaptive algorithm computes a FIR filter that best equalizes these
two paths. These algorithms are effective when the reference
sensors are coherent with the error signals and have a small time
delay with respect to the source, and the system controlled is
linear. The Filtered-X LMS algorithm is described in the textbook
"Adaptive Signal Processing written by Bernard Widrow and Samuel
Stearns.COPYRGT. 1985, Prentice-Hall Inc., ISBN: 0130040290".
[0169] The preferred Fx-LMS algorithm also receives on line 516 the
f.sub.0 signal representing the rotor rotational rate from the
tachometer board 505, as described above. The f.sub.0 signal is
used as a reference by the Fx-LMS in its processing to generate
output values which will minimize the array values received from
the compensation function 321.
[0170] Preferably, the Fx-LMS function also receives a signal on
line 365 from the weight limiting 500 function, described below.
This signal causes the Fx-LMS to freeze its adaptive coefficients
until another signal on line 365 permits the Fx-LMS to continue
adjusting its adaptive coefficients.
The Upshifting Function
[0171] For each output frequency to be controlled (i.e.,
attenuated) (4P, 8P and 12P), the upshifting 360 function receives
from the adaptive filter 325 one complex number for each controlled
output. The upshifting 360 function performs a frequency "upshift"
of the vibratory load signal at base-band to NP, as described
relative to the downshifting function 270 above. An additional
input to this function is the f.sub.0 parameter from the tachometer
process on line 516, from which 4P, in the embodiment shown in FIG.
6, is computed and used to generate a complex sinusoidal look-up
table which performs the 4P frequency up-shift. The complex number
is converted to a single real output signal by the following
calculations
OutputSignal[i][n]=Re{(A+jB)e.sup.j.omega.n}
[0172] Where
[0173] i=1 . . . Number of Controlled Outputs
[0174] Re(X)=real portion of complex number X
[0175] (e.sup.j.omega.n)=cos(.omega.n)+j sin(.omega.n) and
[0176] .omega.=2*.pi.*frequency currently being processed (i.e.
4P/8P/12P in the preferred embodiment)
[0177] n=the current sample count.
[0178] After the upshifting 360 function is applied to the output
of the adaptive filter 325, the real part of the resultant complex
number is extracted. This is indicated as Re { } in FIG. 6B. The
necessary sine and cosine tables are generated as described
relative to the downshifting function 270 above. In the embodiment
shown in FIGS. 6A and 6B, one set of tables is generated and shared
between the downshifting and upshifting functions 270, 360.
[0179] Vibration Post Processing
[0180] After the Vibration Control Phase has generated an output
control value for each controlled output, an additional output
scaling function 370 is applied to the output value received from
the upshifting function 360. The output scaling function 370
permits an additional adjustment of the degree of responsiveness
the controlled device will exhibit to the signals generated by the
Vibration Control Phase. In the embodiment shown in FIGS. 6A and
6B, the output scaling function 370 adjusts the degree of actuation
of the actuators 36 (not shown) used to control vibrations. The
output scaling function 370 consists of a single multiplication of
each upshifting function 360 output by a predetermined real number,
the ValveGainFactor. Alternatively, the output scaling function 370
may apply more involved mathematical functions to the output
control value than the described single multiplication.
[0181] After the output scaling function 370, the weight limiting
function 500 receives the output control signals on line 206 and
makes one final evaluation of the output control values. The weight
limiting function 500 compares the output control values with
threshold values. If the threshold values are exceeded, the weight
limiting function 500 sends a signal on line 365 which causes the
adaptive filter 325 to freeze its adaptive coefficients. When the
output control values subsequently fall below the threshold values,
the weight limiting function 500 sends a signal on line 365 which
allows the adaptive filter 325 to continue adapting its
coefficients.
[0182] The adaptive filter 325 will tend to increase its output
values in efforts to converge to a solution. In some limited
circumstances, these efforts to converge to a solution will result
in output control values which could possibly damage the controlled
device. By freezing the adaptive filter 325 coefficient values when
this objectionable situation occurs, the weight limiting function
500 provides an additional safety measure. In the embodiment shown
in FIGS. 6A and 6B, the weight limiting function 500 incorporates
predetermined threshold values specified so as to prevent the
actuators 36 from exceeding their actuatorial authority and
damaging either the actuators or the attached plant.
[0183] The weight limiting function 500 could optionally be
eliminated from a Controller 100 used in an application in which no
concern over the scale of output signals existed. Such situations
could include attachment to a plant that cannot be damaged by
objectionable output control values or inclusion of measures within
other portions of the Controller 100 to prevent the output control
values transmitted on line 206 from reaching objectionable
levels.
Position Control Processing Path
[0184] Referring to FIG. 6A, after the common position sensor and
vibration sensor scaling and pre-processing function 220, 420, the
position control path processes the signals received from the
position sensors 38. The position control path consists of the
common scaling and pre-processing function 420 (discussed above
with respect to the vibration control processing path), the band
elimination function 440, the position control compensation
function 460, and the output scaling function 480 (FIG. 6B). The
goal of the position control processing path is to keep the
controlled device properly positioned so that the vibration control
path has as much authority as possible to reduce the vibrations at
the controlled frequencies 4P, 8P and 12P while providing static
load to the attached structure.
[0185] For example, in the ATM for a rotary wing aircraft, each
actuator 36 has a position sensor 38 that measures the current
positional displacement of the actuator 36. The goal of position
control path is to keep the actuators properly positioned around a
desired point while providing static loads (vertical actuators) and
thrust/torsional loads (horizontal actuators) to the structure.
[0186] The position control algorithm should be structured to be
flexible and allow for several possibilities. Open loop integral
response is desired for zero steady state error. A fixed, low
bandwidth, near-DC, broadband control compensation is desirable. In
one embodiment of the ATM controller, this corresponds to a first
order low-pass transfer function positional closed loop
response.
[0187] As shown in FIG. 6A, the band elimination function 440,
which in a preferred embodiment is a notch filter, receives the
scaled and preprocessed position sensor signals from the scaling
and pre-processing function. The band elimination function 440
removes any components of the frequencies of interest of the
vibration control signal present in the position control signal.
The position of the controlled device will be changing based on the
signals generated by the vibration control path. The goal of the
position control path is to counteract low frequency movements, but
not those at the frequencies of interest. By removing the
frequencies of interest from the position control signal of the
output control signal, the Controller 100 reduces the possibility
that the position control path will generate signals which
counteract the vibration control efforts of the vibration control
paths. The band elimination function 440 is designed using the
techniques described above for the notch filter of the vibration
controller.
[0188] The band elimination function 440 transmits the resulting
position control signals to the summing box 450. The summing box
450 computes the difference between the desired position of the
actuator and the current scaled, pre-processed and band eliminated
position sensor signal.
[0189] The output of the summing box function 450 is processed by
the position control compensation function 460 using the
proportional/integral/derivative (PID) algorithm, which is well
known in the art. One reference in which the PID algorithm is
described is Analog and Digital Control System Design:
Transfer-Function, State-Space, and Algebraic Methods written by
Chi-Tsong Chen and published by Saunders College Publishing
Electrical Engineering (1995) (ISBN: 0030940702). The following
general relations define the PID compensation scheme adopted for
the position control path:
[0190] The user defined parameters are a set of weighting constants
as well as a nominal, desired position represented as:
[0191] k.sub.1j--direct error weight
j=1,n (n=number of controlled devices)
[0192] k.sub.2j--first difference error weight
j=1,n
[0193] k.sub.3j--second difference error weight
j=1,n
[0194] k.sub.4j--integral error weight
j=1,n
[0195] x.sub.0j--desired position
j=1,n
[0196] The processing calculations are performed as follows:
x.sub.j(k)=The position sensor input at sample k, controlled device
j
e.sub.j(k)=x.sub.0j-x.sub.j(k)=The position error (i.e. offset from
the desired position) at sample k, valve j
es.sub.j(k)=es.sub.j(k-1)+e.sub.j(k)=The accumulated position
error
y.sub.j(k)=k.sub.1je.sub.j(k)+k.sub.2j[e.sub.j(k)-e.sub.j(k-1)]+k.sub.3j[e-
.sub.j(k)-2e.sub.j(k-1)+e.sub.j(k-2)]+k.sub.4jes.sub.j(k)=The
position component of output control signal at sample k for
controlled device j
[0197] The algorithm thus accommodates proportional, integral and
derivative processing options corresponding to zero, first and
second differences. Because the PID user defined constants might be
unique for each controlled device, the position control
compensation function 460 must store independent, unique PID
constants for each controlled device. During each execution frame
of the Controller 100, the position control path is executed once
for each controlled device using the PID constants associated with
the currently processed controlled device.
[0198] The desired position, x.sub.0j, is the point within the
controlled device's range at which maximum authority exists in all
control directions for the vibration controller to utilize in
reducing vibration. In a hydraulic actuator, for example, the
desired position is in the area of the mid-point of the traversing
path of the actuator 36 piston. It is the location which
accommodates both positive and negative (asymmetric) signal swings
for the signals corresponding to the frequencies in interest.
[0199] In the embodiment of the present invention shown in FIGS. 6A
and 6B, the Controller 100 controls two hydraulic control valves 48
(FIG. 6B). In a rotary wing aircraft, each valve 48 could supply
hydraulic fluid to a single actuator of an ATM mount 32, including
a vertical actuator and a horizontal actuator. The desired position
of the actuator piston is empirically determined by the user and is
based upon the optimal position that minimizes enhancement of the
frequencies of interest, given the fact that the vertical load is
much higher than the horizontal load because the weight of the
attached plant, in this case the airframe, is usually on the
vertical load. This corresponds to a position off the mid-point of
the vertical actuator 36 piston. The horizontal actuator 36 remains
close to its mid-point trajectory. In the rotary wing aircraft, the
desired actuator 36 position can be determined automatically, for
example, by monitoring the maximum extents that occur during
operation of the Controller 100 and altering the desired position
to be the midpoints of those extents.
[0200] The values of the user-defined coefficients depend on the
actuator valve to position sensor transfer function. The actuator
valve to position sensor transfer function is determined by the
characteristics of the controlled devices. For example, if the
controlled device is a hydraulic actuator, multiple methods of
actuation are known in the art and some of these methods result in
different transfer functions. Two such differing actuation systems
are flow control and pressure control. Direct command emphasizing
the direct error weight, k.sub.1j, would usually be used for flow
control actuation systems while derivative control emphasizing the
integral weight, k.sub.4j, would be used for pressure control
actuation systems.
[0201] In an active control system, the hydraulic valve 48 used for
the ATM actuators 36 is preferably a flow control valve. In such a
system, the low-frequency response of position to a flow command
would be a single integral which calls for a direct command. It has
been verified through measurements that the resultant actuator
valve to position sensor transfer function approaches that of an
integrator in the frequencies of interest. This measured gain is of
the simplified form at low frequencies (neglecting the small phase
delay) 5 P ( s ) = K 0 s
[0202] where P(s) is the actuator valve to position sensor transfer
function, K.sub.0 is a constant and s is the Laplacian variable.
This actuator valve to position sensor transfer function is defined
independently of the PID coefficient settings. However, the optimal
PID coefficient settings are dependent on the transfer function,
especially since there is a small (but non-zero) phase delay in the
position control path attributed to the A/D and D/A process. This
actuator valve to position sensor transfer function above requires
nonzero coefficients for the direct terms, k.sub.1j, and zero for
the difference and integral components, k.sub.2j, k.sub.3j, and
k.sub.4j. The direct coefficients are determined analytically in
order to prevent position controller instability. Methods of
determining the coefficients for a particular plant transfer
function are well known in the art. Equations and control design
techniques are discussed in the Chen reference cited above.
[0203] Referring to FIG. 6B, after the position control path has
generated an output position control value for each controlled
device, an additional output scaling function 480 is applied to the
output value position control compensation function 460. As
described above in the vibration control post processing, the
output scaling function 480 permits an additional adjustment of the
degree of responsiveness by the controlled device. In a preferred
embodiment, the output scaling function 480 consists of a single
multiplication of each position control compensation function 460
output by a predetermined real number, the ValveGainFactor.
Position Control Pre-Run (ATMPC Pre-Run)
[0204] The ATMPC pre-run 410 function ascertains the initial
average control static signal required to hold an average position.
By determining the actuator valve 48 to position sensor 38 transfer
function, the ATMPC pre-run 410 function permits the Controller 100
or user to more readily determine coefficients for the position
control algorithm which, in a preferred embodiment, is the PID
system described above. The pre-run function 410 is executed during
a system initialization or setup phase. Sample output control
values are sent to the controlled device and the resulting effect
upon the input position sensor signals are correlated with the
sample output control values. Based on this correlation,
coefficients for the position control algorithm are calculated.
During pre-run 410, either the raw output control value/input
sensor signals or the calculated coefficients are stored to a
physical medium like a disk drive 415 (FIG. 6A). An offline
evaluation could be performed upon the raw data with the resulting
coefficients programmed directly into the position compensation
function 460 described above.
[0205] In the embodiment shown in FIGS. 6A and 6B, the ATMPC
pre-run 410 function is implemented as follows:
[0206] x=valve command
[0207] x.sub.0=valve command for zero actuator motion
[0208] x.sub.01=estimate of valve command at minimum actuator
position, y.sub.1
[0209] x.sub.02=estimate of valve command at maximum actuator
position, y.sub.h
[0210] x.sub.h=maximum allowable valve command-input to this
procedure
[0211] x.sub.1=minimum valve command-input to this procedure
[0212] y=actuator position indication
[0213] y.sub.h=maximum actuator position
[0214] y.sub.1=minimum actuator position
[0215] y.sub.0=actuator center position=(y.sub.h+y.sub.1)/2
[0216] K=actuator system gain=actuator speed/valve command
[0217] T=period of applied square wave
[0218] dx=amplitude of applied square wave
[0219] The objective is to determine for each actuator/valve
combination the values of x.sub.0, y.sub.0, and K. First, the ATMPC
pre-run 410 function determines x.sub.01 and y.sub.1 as follows.
Starting with the minimum valve voltage level, x.sub.1, the
corresponding actuator position, y.sub.1, is measured. The valve
command voltage at which the actuator position begins to move is
x.sub.01. Second, x.sub.02 and y.sub.h are determined as follows.
Starting with the maximum valve voltage, x.sub.h, the corresponding
actuator position, y.sub.h, is measured. The valve command at which
the value of y begins to decrease is x.sub.02. Third, actuator
midpoint, y.sub.0=(y.sub.h+y.sub.1)/2, and the valve command for
average center position, x.sub.0=(x.sub.01+x.sub.02)/2, are
calculated. Fourth, K is measured K which, in the embodiment shown
in FIGS. 6A and 6B, is the actuator speed for a flow control valve
48. Beginning with the minimum valve command, x.sub.1, a square
wave between x.sub.0 and x.sub.0+dx with period T is applied until
the actuator position reaches y.sub.0. Next, a square wave between
x.sub.0-dx and x.sub.0+dx with period T is applied. The output
actuator position forms a triangular wave with amplitude dy. The
system gain K can be calculated with the formula K=dy/(Tdx). This
K, thus, corresponds to the transfer function of an integrator of
the form: 6 F ( s ) = Y d ( s ) X d ( s ) = K s
[0220] where Y.sub.d(s) corresponds to the Laplace transform of the
signal dy and X.sub.d(s) corresponds to the Laplace transform of
the signal dx.
[0221] Preferably, the ATMPC pre-run 410 function is executed
during initial system installation. Since the actuator valve to
position sensor transfer function does not change frequently for a
rotary wing aircraft application, the ATMPC pre-run 410 function
would normally only be run after substantial changes have occurred
in the associated plant. Preferably, for a rotary wing application,
the position control prerun function would be performed
automatically on system initialization.
Control Signal Summation
[0222] The summing function 62 (FIG. 6B) receives the vibration
control output signals on line 206 from the vibration control path
and the position control output signals on line 228 from the
position control path and combines these signals. Preferably, the
combination of signals is accomplished through an arithmetical
addition of digital values. Other means of combining the signals
include using a weighted function wherein the control signals input
to the summing 62 function are given different relative weights in
the resultant output signals. Further mathematical functions beyond
addition could also be applied to the input control signals.
[0223] After combination, the combined signals are output to the
hydraulic servo valves 48 on line 52. Because a hydraulic control
valve 48 may respond to analog signals only, a conversion from
digital format to analog format by a D/A converter (not shown) may
be necessary before sending the combined signal to the hydraulic
control valve 48. The D/A converter receives the digital combined
signal on control line 52, converts that signal to analog form
appropriate for the attached plant and then transmits that analog
signal to the controlled device on a signal line which runs from
the D/A converter to the controlled device.
[0224] The previously described embodiments of the present
invention have many advantages, including effective control for an
active system for reducing the transmission of vibration and noise
between a vibrating component and a structure. The controller of
the present invention transmits at least two output vibration
cancellation signals which control a vibration cancellation
mechanism while ensuring that sufficient authority exists in the
vibration cancellation mechanism to respond to the vibration
cancellation signals. The control system is particularly effective
when the vibration cancellation mechanism is located within the
connection points, or in series, between a vibration generating
component and the mounting location of the component on the
structure. In a rotary wing aircraft application including
actuators mounted between the transmission gearbox feet and
airframe, the controller for the active system isolates the main
rotor system of the helicopter, and prevents the low frequency
vibration generated by the main rotor system from being transmitted
to the airframe while efficiently passing the primary flight loads.
This active system must also maintain the average static position
of the transmission relative to the airframe. The controller
functions to maintain the actuator position at a predetermined
point within a maximum range to which the vibration producing means
is capable of responding to ensure the actuator has sufficient
authority to respond to the vibration cancellation signals.
Moreover, the effectiveness of passive vibration reduction
techniques are enhanced when used in conjunction with the active
vibration control system of the present invention.
[0225] Although the present invention has been shown and described
in considerable detail with respect to only a few exemplary
embodiments thereof, it should be understood by those skilled in
the art that I do not intend to limit the invention to the
embodiments since various modifications, omissions and additions
may be made to the disclosed embodiments without materially
departing from the novel teachings and advantages of the invention,
particularly in light of the foregoing teachings. For example, it
is apparent that the present invention could also be embodied in
hardware circuitry which performs the same functional operations.
Accordingly, I intend to cover all such modifications, omission,
additions and equivalents as may be included within the spirit and
scope of the invention as defined by the following claims. In the
claims, means-plus-function clauses are intended to cover the
structures described herein as performing the recited function and
not only structural equivalents but also equivalent structures.
Thus, although a nail and a screw may not be structural equivalents
in that a nail employs a cylindrical surface to secure wooden parts
together, whereas a screw employs a helical surface, in the
environment of fastening wooden parts, a nail and a crew may be
equivalent structures.
* * * * *