CN107792753B

CN107792753B - Active vibration damping device for elevator

Info

Publication number: CN107792753B
Application number: CN201710522673.7A
Authority: CN
Inventors: 平井正昭; 藤原琢也
Original assignee: Toshiba Elevator Co Ltd
Current assignee: Toshiba Elevator and Building Systems Corp
Priority date: 2016-09-05
Filing date: 2017-06-30
Publication date: 2019-12-31
Anticipated expiration: 2037-06-30
Also published as: JP2018039591A; CN107792753A; JP6242969B1

Abstract

An active vibration damping device for an elevator, comprising: a cage that moves up and down along the guide rail; a vibration damping mechanism provided at a portion of the car facing the guide rail; at least one displacement sensor for detecting a relative displacement between the car and the guide rail; an estimation means having a mathematical model theoretically representing a relative displacement between the car and the guide rail due to a deflection amount of the guide rail, for estimating the deflection amount of the guide rail, which causes vibration of the car during traveling, substantially in real time using a signal of the displacement sensor and the mathematical model; and a control unit that controls the vibration reduction mechanism in a direction in which vibration of the car is suppressed, based on a result of the estimation by the estimation unit.

Description

Active vibration damping device for elevator

Technical Field

Embodiments of the present invention relate to an active vibration damping device for an elevator, which suppresses vibration during car travel.

Background

As elevators are accelerated, the importance of techniques for suppressing vibration (horizontal vibration) during car travel is increasing. The largest cause of the vibration generated during running is a minute deflection of the guide rail (guide rail). That is, when the car travels along the guide rail, if the car travels at a high speed on the guide rail which originally has a slight deflection, the car is forcibly displaced and vibrates in the horizontal direction.

Therefore, an active roller guide (active roller guide) technique for actively suppressing vibration of the car has been proposed. The active guide shoe is disposed to abut against the guide rail at the upper and lower sides of the car. An acceleration sensor is provided in the car, and the vibration of the car is actively suppressed by feedback-controlling an actuator of the guide shoe based on the vibration detected by the acceleration sensor during running.

The vibration (horizontal vibration) generated during the running of the elevator is, for example, micro vibration with a frequency of about 1 to 5Hz and an amplitude of about 0.02 to 0.03G, and is much smaller than the vibration of vehicles such as railways and buses.

Here, a low-cost MEMS acceleration sensor, which is generally used as a vibration sensor, is designed to detect vibrations of about 1 to 2G. Therefore, the MEMS acceleration sensor cannot accurately detect the micro vibration of the elevator, and it is difficult to perform vibration damping control using the MEMS acceleration sensor. Therefore, it is necessary to use a high-precision acceleration sensor such as a servo-type acceleration sensor in the vibration damping system of the elevator.

However, such an acceleration sensor is expensive, and when a plurality of acceleration sensors are used to improve accuracy, the cost of the vibration damping system is significantly increased, which is not practical. Therefore, the active shoe described above is required to perform vibration damping control without using an acceleration sensor.

Disclosure of Invention

The problem to be solved by the present invention is to provide an active vibration damping device for an elevator, which can reliably capture and effectively reduce vibration caused by deflection of a guide rail during car traveling by an inexpensive structure.

An embodiment provides an active vibration damping device of an elevator, which is characterized by comprising:

a cage that moves up and down along the guide rail;

a vibration damping mechanism provided at a portion of the car facing the guide rail;

at least one displacement sensor for detecting a relative displacement between the car and the guide rail;

an estimation means having a mathematical model theoretically representing a relative displacement between the car and the guide rail due to a deflection amount of the guide rail, for estimating the deflection amount of the guide rail, which causes vibration of the car during traveling, substantially in real time using a signal of the displacement sensor and the mathematical model; and

and a control unit that controls the vibration reduction mechanism in a direction in which vibration of the car is suppressed, based on a result of the estimation by the estimation unit.

According to the embodiment, the active vibration damping device of the elevator can be provided, which can reliably capture and effectively reduce the vibration caused by the deflection of the guide rail during the running of the car through an inexpensive structure.

Drawings

Fig. 1 is a diagram schematically showing the structure of an active vibration damping device for an elevator according to embodiment 1.

Fig. 2 is a diagram showing a structure of a drive shoe provided in a car according to this embodiment.

Fig. 3 is a diagram showing a configuration of a control system of the active shoe according to the embodiment.

Fig. 4 is a diagram for explaining a two-degree-of-freedom vibration system model according to this embodiment.

Fig. 5 is a diagram schematically showing a two-degree-of-freedom vibration system model according to this embodiment.

Fig. 6 is a diagram for explaining the structure of the scope of the embodiment.

Fig. 7 is a diagram for explaining another configuration of the scope according to the embodiment.

Fig. 8 is a block diagram showing a functional configuration of the control device according to the embodiment.

Fig. 9 shows a state in which the active shoe of this embodiment is subjected to forced displacement due to deflection of the guide rail, fig. 9(a) shows a state before displacement, and fig. 9(b) shows a state after displacement.

Fig. 10 is a diagram schematically showing the processing of the feedforward control module according to this embodiment.

Fig. 11 is a diagram for explaining the characteristics of the bent state of the guide rail according to this embodiment.

Fig. 12 is a diagram showing a relationship between an amplitude component and a period of a deflection waveform of the guide rail according to this embodiment.

Fig. 13 is a diagram showing the comparison between the amount of deflection of the guide rail according to the present embodiment and the estimation result.

Fig. 14 is a diagram showing the vibration of the car and the result of suppressing the vibration in comparison in the embodiment.

Fig. 15 is a block diagram showing a functional configuration of the control device according to embodiment 2.

Fig. 16 is a diagram schematically showing the structure of an active vibration damping device for an elevator according to embodiment 3.

Fig. 17 is a diagram schematically showing the structure of an active vibration damping device for an elevator according to embodiment 4.

Fig. 18 is a diagram for explaining the structure of the scope according to embodiment 5.

Description of the reference symbols

1 … well; 2-1, 2-2 … guide rails; 3 … bracket; 4 … ropes; 5 … car; 6 … car frame; 7-1 to 7-4 … active guide shoes; 8-1 to 8-4 … guide wheels; 9-1 to 9-4 … supporting members; 10-1 to 10-4 … springs; 11-1 to 11-4 … actuator; 12 … car room; 13-1, 13-2 … vibration-proof rubber; 14-1, 14-2 … actuators; 15-1 to 15-6 … displacement sensors; 16-1 to 16-4 … relative shift signal; 20 … control device; 21-1, 21-2 … driving device; 31 … rail displacement estimation module; 32-1 to 32-4 … guide rail displacement estimation signals; 33 … a feed forward control module; 34-1 to 34-4 … feed forward control signals; 35 … feedback control module; 36-1 to 36-4 … feedback control signals; 37-1 to 37-4 … adder; 38-1 to 38-4 … vibration control signals; 40 … model; 41. 61 … observer (estimator).

Detailed Description

Hereinafter, embodiments will be described with reference to the drawings.

(embodiment 1)

A pair of guide rails 2-1 and 2-2 are vertically arranged in an elevator shaft 1 (lifting road). The guide rails 2-1, 2-2 are fixed by a plurality of brackets (blackets) 3 disposed at equal intervals in the vertical direction on the wall surface of the hoistway 1. The car 5 is supported to be freely raised and lowered on the guide rails 2-1, 2-2. The car 5 is driven by a hoisting machine, not shown, to move up and down in the hoistway 1 via the rope 4.

Here, active guide shoes 7-1 to 7-4 are provided at 4 positions of a car frame 6 constituting an outer frame of a car 5. The active guide shoes 7-1 to 7-4 actively reduce lateral vibration generated in the car 5 and guide the running. The active guide shoes 7-1 and 7-2 are provided on the upper and lower portions of one side (right side in the drawing) of the car frame 6, and abut against one of the guide rails 2-1. The active guide shoes 7-3 and 7-4 are provided on the upper and lower portions of the other side (left side in the drawing) of the car frame 6, and abut against the other guide rail 2-2.

Further, displacement sensors 15-1 and 15-2 for detecting relative displacement between the car 5 and one of the guide rails 2-1 are provided at respective positions of the active guide shoes 7-1 and 7-2. Similarly, displacement sensors 15-3 and 15-4 for detecting relative displacement between the car 5 and the other guide rail 2-2 are provided at the respective installation positions of the active shoes 7-3 and 7-4.

The displacement sensors 15-1 to 15-4 are non-contact displacement sensors. Examples of the detection method include an eddy current type, a capacitance type, an ultrasonic type, and an optical type, but the present invention is not particularly limited to these methods.

Fig. 2 is a diagram showing a structure of the active shoe 7-1 provided in the car 5. Here, the structure of the active guide shoe 7-1 provided on the upper right side of the car frame 6 is shown, but the other active guide shoes 7-2 to 7-4 have the same structure.

The active shoe 7-1 is provided with a guide wheel 8-1 abutting against the guide rail 2-1, a support member 9-1 supporting the guide wheel 8-1, and a spring 10-1 pressing the guide wheel 8-1 against the guide rail 2-1. In reality, there are a total of 3 guide wheels including 2 front and rear direction wheels sandwiching the guide rail 2-1 from the front and rear direction for guiding the car in the front and rear direction, but only 1 guide wheel for guiding and supporting the car in the left and right direction is shown here.

The active shoe 7-1 includes an actuator (activator) 11-1 for vibration reduction in addition to these general guide mechanisms. The actuator 11-1 is disposed between the car 5 and the support member 9-1, and generates an arbitrary force between the guide wheels 8-1 and the car 5 in addition to the pressing force of the spring 10-1.

Here, the displacement sensor 15-1 is disposed in the vicinity of the support member 9-1. In detail, as shown in fig. 2, the displacement sensor 15-1 is fixed to a fixed member 6a extending from the car frame 6, and detects a distance d between the support member 9-1 and the car frame 6. The distance d represents the relative displacement between the car 5 and the guide rail 2-1 at the location where the active shoe 7-1 is disposed.

The same applies to the other displacement sensors 15-2 to 15-4. In the present embodiment, these displacement sensors 15-1 to 15-4 are used to detect the relative displacement in the left-right direction (x direction) between the car 5 and the guide rails 2-1, 2-2 at 4 locations of the car 5, thereby reducing the vibration in the left-right direction (x direction) of the car 5.

The vibration of the car 5 in the front-rear direction (y direction) can be reduced by the same method. In this case, the displacement sensors 15-1 to 15-4 are provided to detect relative displacement in the front-rear direction (y direction) between the car 5 and the guide rails 2-1 and 2-2, respectively. The system for reducing the vibration in the left-right direction and the system for reducing the vibration in the front-rear direction can be provided simultaneously.

Further, the reason why the displacement sensors are provided at the upper and lower portions of the car 5 is that it is assumed that the car 5 is a two-degree-of-freedom vibration system having horizontal vibration and rotational vibration. In the model of the two-degree-of-freedom vibration system, at least a displacement sensor for detecting a horizontal vibration component of the car 5 and a displacement sensor for detecting a rotational vibration component of the car 5 are required. When the car 5 is modeled on the one-degree-of-freedom vibration system, the displacement sensor may be provided at one position of the upper part or the lower part of the car 5.

Hereinafter, a method of reducing the vibration in the left-right direction (horizontal vibration) assuming that the car 5 is a two-degree-of-freedom vibration system will be described.

In general, the guide rails 2-1 and 2-2 are formed by connecting a plurality of guide rail members having a predetermined length in a vertical direction. It is extremely difficult to erect the guide rails 2-1, 2-2 completely vertically, and there is a slight deflection (bending) at the stage of installation. The minute flexure (bending) acts as a forced displacement when the car 5 travels, and generates a horizontal shake (horizontal vibration). In order to actively suppress such horizontal vibration, actuators 11-1 to 11-4 as vibration damping means are provided to the active shoes 7-1 to 7-4.

Fig. 3 is a diagram showing an example of the configuration of the control system of the active shoes 7-1 and 7-2. In fig. 3, the control system is shown for the active shoes 7-1 and 7-2 provided on the upper right side and the lower right side of the car frame 6, but the same applies to the other active shoes 7-3 and 7-4.

The signals of the displacement sensors 15-1, 15-2 provided to the active shoes 7-1, 7-2 are input to the control device 20. Here, when the signal of the displacement sensor 15 is an analog signal, the signal is input to the control device 20 via an a/D converter not shown. On the other hand, when the signal from the displacement sensor is a digital signal, the signal is directly input to the control device 20 by wired or wireless communication.

The control device 20 includes a microcomputer, and is provided in the car 5. The control device 20 executes arithmetic processing for reducing the vibration of the car 5 at predetermined cycles (for example, 1ms cycles) based on the signals of the displacement sensors 15-1 and 15-2.

The driving devices 21-1 and 21-2 are provided in the car 5, and drive the actuators 11-1 and 11-2 in accordance with a drive control signal (force command signal or displacement command signal) output from the control device 20. In practice, driving means corresponding to the actuators 11-3 and 11-4 are provided, and the actuators 11-3 and 11-4 are driven in accordance with a drive control signal (force command signal or displacement command signal) output from the control device 20. Thus, when the car 5 vibrates in the horizontal direction, the actuators 11-1 to 11-4 work in a direction to suppress the vibration.

Horizontal vibration (lateral vibration) of an elevator includes vibration in the left-right direction (x direction) and vibration in the front-back direction (y direction). Hereinafter, the vibration in the left-right direction will be described as an object, but the vibration in the front-back direction can be similarly applied.

Here, in the present embodiment, the deflection amount of the guide rails 2-1, 2-2 causing the vibration of the car 5 is estimated in real time by an observer (estimator) during traveling, and the actuators 11-1 to 11-4 generate feedforward control forces for canceling the vibration caused by the deflection amount, thereby suppressing the vibration. As a method for realizing this control, a mathematical model that theoretically shows the relationship between the deflection of the guide rails 2-1, 2-2 and the horizontal vibration (left-right vibration) of the car 5 is constructed in the control device 20.

The mathematical model includes an "extended equation of state model" obtained by combining a "car vibration model" and a "guide rail displacement model".

"car vibration model" is a model that expresses the vibration characteristics of the car 5 when the car is forcibly displaced in the horizontal direction by the deflection of the guide rails 2-1, 2-2 in the form of a state equation (see expressions (1) and (2), and expressions (4) and (5)).

"guide rail displacement model" is a model that is assumed to express the deflection of the guide rails 2-1, 2-2 in the form of a state equation with the deflection changing according to a predetermined regular characteristic (see equations (6) and (7)).

Further, "guide rail displacement" is displacement in the horizontal direction due to flexure (bending) of the guide rails 2-1, 2-2, and is zero in a state where the guide rails 2-1, 2-2 are vertically disposed straight without flexure in the vertical direction. This rail displacement is also sometimes referred to as "deflection".

The "extended equation of state model" is an equation of state obtained by combining the "car vibration model" and the "guide rail displacement model" (see equations (8) and (9)).

The mathematical model is explained in detail.

'Car vibration model'

As shown in fig. 4, a two-degree-of-freedom vibration system model having two vibration angles θ (t) around the center of gravity and a horizontal vibration shift x (t) of the center of gravity is considered as the vibration characteristics of the car 5. Schematically representing the two-degree-of-freedom vibration system model is shown in fig. 5.

The U1, U2 in the figure are the forces of the actuators 11-1, 11-2 in the upper and lower car parts. In addition, as specifications of the car 5, the car weight is m (kg), and the inertia moment is J (kg · m)²) Distances from the center of gravity to the upper and lower guide shoes are set to L1 and L2[ m ]]The spring constant of the upper and lower guide shoes is set to K [ N/m ]]The damping constant of the spring is C [ Ns/m ]]。

When the horizontal vibration shift of the center of gravity of the car 5 is represented by X (t), the rotational vibration angle around the center of gravity is represented by θ (t), and the time differentials of these are represented by X '(t) and θ' (t), the state equations of the two-degree-of-freedom vibration system model are generally expressed by the following expressions (1) and (2).

The equations (1) and (2) are a set of equations, which are the running equation and the output equation. The point above the symbol represents a 1 st order differential and the two points represent a 2 nd order differential. The equation (2) is a case where the vibration displacements X + L1 θ and X-L2 θ near the upper and lower shoes are detected by the acceleration sensor as shown in fig. 5.

Here, a [4 × 4] and B [4 × 4] are matrix-type constant matrix values uniquely determined by specification values of the car 5. Since this is the case, a description of specific values is omitted. The specifications of the car 5 include, for example, car weight, inertia moment, distance from the center of gravity to the upper and lower guide shoes, spring constants of the upper and lower guide shoes, damping constants included in the spring, and the like. C4 x 4 is a matrix form which summarizes the previous items, and a specific formula is omitted.

When the forced displacement vectors D1, D2, D1', and D2' are given as the disturbances, what kind of horizontal vibration displacement x (t) and rotational vibration angle θ (t) are generated in the car 5, and what kind of signals are output from acceleration sensors (not shown) provided in the upper and lower portions of the car 5, can theoretically be calculated by the above equations (1) and (2).

However, in the present embodiment, since the shift sensor is used instead of the acceleration sensor, it is necessary to modify a mathematical model capable of calculating the output Y from the signal of the shift sensor.

In fig. 3, the relative displacement between the car 5 and the guide rail 2-1 detected by the displacement sensor 15-1 is h1, the relative displacement between the car 5 and the guide rail 2-1 detected by the displacement sensor 15-2 is h2, the guide rail displacement (amount of deflection) of the guide rail 2-1 at the installation location of the upper active shoe 7-1 is D1, and the guide rail displacement (amount of deflection) of the guide rail 2-1 at the installation location of the lower active shoe 7-2 is D2. The following formula (3) can be obtained as a relational expression of these geometries.

X+L1θ-D1＝h1

X-L2θ-D2＝h2...(3)

Here, (X + L1 θ), (X-L2 θ) represent absolute shifts. That is, on the upper shoe side, a value obtained by subtracting the shift D1 from a value obtained by adding the rotational vibration shift L1 θ to the horizontal vibration shift X is obtained as the relative shift h 1. Similarly, on the lower shoe side, a value obtained by subtracting the displacement D2 from the value obtained by adding the rotational vibration displacement-L2 θ to the horizontal vibration displacement X is obtained as the relative displacement h 2.

When the state equations of the above equations (1) and (2) are rewritten using the equation (3), the equations (4) and (5) are obtained.

Equations (4) and (5) are a set of equations, a running equation and an output equation. Equation (4) is an operating equation including interference (shift) as in equation (1). Equation (5) is an equation in which equation (2) is changed to a form in which a signal of the displacement sensor can be used. The output Y is the estimated amount of relative shift. The relative displacement speed and the relative displacement on the upper shoe side are represented by (h1', h1), and the relative displacement speed and the relative displacement on the lower shoe side are represented by (h 2', h 2).

'guide rail shift model'

Here, the guide rail displacement model can be modeled under a certain assumption, focusing on the installation environment of the guide rails 2-1 and 2-2. That is, the guide rails 2-1 and 2-2 are fixed to a wall surface of the hoistway 1 by a bracket 3 (see fig. 11) as a fixing member. Since the flexure (bending) is suppressed at the fixed point, the flexure tends to occur at a cycle starting from the fixed point. Therefore, it is assumed that the deflection of the guide rails 2-1 and 2-2 changes and is modeled according to a substantially sinusoidal wave characteristic having a period at the installation interval of the carriage 3.

For example, assuming a sine wave of the frequency ω of the vibration, the frequency is expressed by equations (6) and (7).

"extended equation of state model"

When an "extended equation of state model" is constructed by combining the above equations (4) to (7), equations (8) and (9) are obtained.

Equations (8) and (9) are a set of equations, which are an operation equation and an output equation in an extended state equation model (car vibration + guide rail displacement model) of a two-degree-of-freedom vibration system.

Here, the output equation of the equation (9) is the same as the output equation of the above (5). In the output equation, since the matching between the signal of the displacement sensor, which is an actual measurement value, and the extended state equation is obtained, a conversion equation (part (c) of the equation) for calculating the relative displacement between the guide rail and the car based on the relationship between the state quantity (X ', X, θ', θ) of the car vibration and the estimated quantity (D1', D2', D1, D2) of the guide rail displacement is included.

By configuring an observer using such an extended state equation as a mathematical model and feeding back a difference between a measurement result of the displacement sensor and the output Y of the above equation (9) by multiplying an observer gain (observer gain), it is possible to obtain a rail displacement (amount of deflection) to be estimated.

Fig. 6 is a diagram for explaining the structure of the scope of the present embodiment. In the figure, reference numeral 40 shows a real object (an elevator car) representing an actual vibration system. 41 is an observer. The observer 41 is software, and the real object 40 and the signal of the displacement sensor are hardware. The real object 40 is the car 5, and signals indicating the displacement sensors 15-1 and 15-2 provided to the real object 40 are input to the observer 41.

Now, a case of estimating the displacement of the guide rail due to the deflection of one guide rail 2-1 will be described as an example. Signals indicating the relative displacement between the guide rail 2-1 and the car 5 are input from the displacement sensors 15-1, 15-2 to the observer 41 while the car 5 is traveling.

The observer 41 is composed of a car vibration + guide rail displacement model 42, a relative displacement conversion matrix 43, a difference calculation unit 44, and an observer gain matrix 45.

The car vibration + guide rail displacement model 42 corresponds to the extended state equation expressed by the above equations (7) and (8), and outputs the state quantities (X ', X, θ', θ) of the car vibration and the estimated quantities (D1', D2', D1, D2) of the guide rail displacement. In the figure, (a) to (d) correspond to the portions of the output equations (a) to (d) given by the above expression (9).

The relative shift transformation matrix 43 corresponds to the part (c) of the above equation (9). The relative displacement transformation matrix 43 is designed to derive the relative displacement between the guide rail and the car based on the state quantity (X ', X, θ', θ) of the car vibration and the estimated quantity (D1', D2', D1, D2) of the guide rail displacement.

The relative shift derived from the relative shift transform matrix 43 is an estimate. The difference calculation unit 44 compares the estimated value with the actual measurement value of the relative displacement, and feeds back the difference value to the car vibration + guide rail displacement model 42 via the observer gain matrix 45. The observer gain matrix 45 is a matrix for multiplying a difference value between an estimated value and an actually measured value of the relative shift by a predetermined gain. The observer gain matrix 45 is designed to be able to be larger than the frequency ω of the deflection waveform of the guide rail described later₁The amount of deflection is quickly estimated by the observer 41.

When the observer 41 having such a structure is incorporated in the control device 20 and signals from the displacement sensors 15-1 and 15-2 are input to the control device 20 as shown in fig. 3, the amount of deflection of the guide rail 2-1 can be estimated. When the actuator 11-1 of the active shoe 7-1 and the actuator 11-2 of the active shoe 7-2 are driven by feedforward control based on the estimation result, displacement due to deflection of the guide rail 2-1 can be absorbed in real time, and horizontal vibration of the car 5 can be reduced.

In the configuration of fig. 6, the signals of the shift sensors 15-1 and 15-2 are input to the observer 41 in the original sizes, but the signals of the shift sensors 15-1 and 15-2 may be input after being multiplied by the correction gain 46, as shown in fig. 7, for example.

The correction gain 46 is a gain for improving the estimation accuracy. In general, the signals at the displacement sensors 15-1, 15-2 and the estimate of the observer 41 will be subject to errors due to the modeling of the mathematical model 42. The observer 41 plays a role of suppressing the influence of the model error to be small. In this case, when the correction gain 46 is designed to multiply the sensor signal by, for example, a gain of about 2 to 4 times, the influence of the model error can be strongly corrected, and the rail displacement can be estimated with higher accuracy.

Here, for the sake of simplicity of explanation, the case where the amount of deflection of one guide rail 2-1 is estimated is explained, but actually, the estimation of the amount of deflection of the other guide rail 2-2 by using the displacement sensors 15-3 and 15-4 is also included.

In short, the observer 41 estimates the amount of deflection of the guide rails 2-1 and 2-2 in substantially real time by using an extended equation of state model, using the relative displacement between the guide rails and the car and the relative displacement speed obtained from the signals of the displacement sensors 15-1 to 15-4 as the state quantities of the car 5 as input signals. By performing feedforward control of the vibration damping mechanisms (actuators 11-1 to 11-4) of the active shoes 7-1 to 7-4 based on the estimation result of the observer 41, it is possible to obtain a vibration damping effect similar to that of a mechanism for learning the deflection amount of the guide rails 2-1 and 2-2 in advance.

Hereinafter, a specific configuration will be described.

Fig. 8 is a block diagram showing a functional configuration of the control device 20.

The control device 20 includes a guide rail displacement estimation module 31, a feedforward control module 33, and a feedback control module 35 as functions for suppressing horizontal vibration of the car 5. The relative shift signals 16-1 to 16-4 outputted from the shift sensors 15-1 to 15-4 are supplied to the rail shift estimation module 31 and to the feedback control module 35.

Further, rail displacement estimation signals 32-1 to 32-4 outputted from a rail displacement estimation module 31 described later are supplied to a feedback control module 35.

The feedback control module 35 performs predetermined arithmetic processing using the relative shift signals 16-1 to 16-4 and the rail shift estimation signals 32-1 to 32-4. As a calculation method, for example, when the guide rail displacement estimation signals 32-1 to 32-4 are subtracted from the relative displacement signals 16-1 to 16-4, the vibration displacements of the upper and lower guide shoe portions of the car 5 can be calculated. The vibration shift is further subjected to time differentiation to convert the vibration shift into a vibration velocity, and a value obtained by multiplying the value by a predetermined gain is output as feedback control signals 36-1 to 36-4. In this case, the feedback control force acts as a vibration damping force, and an effect of quickly damping vibration when the car 5 vibrates and an effect of stabilizing the entire control can be expected.

Here, in the present embodiment, the control device 20 includes a rail displacement estimation module 31 and a feedforward control module 33 in addition to the feedback control module 35.

The rail displacement estimating module 31 corresponds to the above-described observer. The guide rail displacement estimation module 31 has a mathematical model representing the characteristics of the horizontal vibration received by the car 5 due to the deflection of the guide rails 2-1 and 2-2, and estimates the deflection amounts of the guide rails 2-1 and 2-2 substantially in real time during running using the relative displacement signals 16-1 to 16-4 and the mathematical model.

The feedforward control module 33 performs predetermined arithmetic processing based on the rail displacement estimation signals 32-1 to 32-4 output from the rail displacement estimation module 31 (see fig. 10).

In addition, the feedback control signals 36-1 to 36-4 correspond to feedback forces generated at the actuators 11-1 to 11-4. The feedforward control signals 34-1 to 34-4 correspond to the actuators 11-1 to 11-4 of the active shoes 7-1 to 7-4, respectively.

The feedforward control signals 34-1 to 34-4 are signals for driving the actuators 11-1 to 11-4 in directions to cancel the forced displacement and the displacement speed due to the deflection of the guide rails 2-1 and 2-2. The details will be described later.

Finally, the feedback control signals 36-1 to 36-4 and the feedforward control signals 34-1 to 34-4 are added by the addition operators 37-1 to 37-4, respectively, to obtain the vibration control signals 38-1 to 38-4. The vibration control signals 38-1 to 38-4 are supplied to the driving devices 21-1 and 21-2 … shown in FIG. 3, and drive the actuators 11-1 to 11-4 of the active shoes 7-1 to 7-4, respectively.

In addition, the vibration can be reduced to nearly 0 in principle by the feedforward control signals 34-1 to 34-4 alone. However, in practice, since the rail displacement estimation signals 32-1 to 32-4 naturally have errors, if the estimation error is large, it is considered that the vibration control may act on the vibration application side and the control may diverge. In such a case, the presence of the feedback control signals 36-1 to 36-4 serves as a component for damping the car vibration and stabilizing the control, so that the influence of the error of the feedforward control signal 44 can be reduced.

Next, the arithmetic processing by the feedforward control module 33 will be described.

Fig. 9 is a diagram showing a state where the driving shoe 7-1 is subjected to forced displacement due to deflection of the guide rail, and fig. 9(a) shows a state before the displacement and fig. 9(b) shows a state after the displacement.

Let the upper active shoe 7-1 for example be subjected to a displacement D1 (t). At this time, assuming that the horizontal position of the car 5 is not changed, the spring 10-1 is displaced to D1[ m ].

Here, the force fr (N) applied to the car 5 by the spring 10-1 is shown below, where K [ N/m ] is the spring constant of the spring 10-1 and C [ Ns/m ] is the damping constant.

Fr＝K·D1(t)+C×D1'(t)

This Fr becomes the excitation force of the car 5.

In contrast, when the actuator 11-1 generates the force of-Fa, the excitation force Fc transmitted to the car 5 becomes Fc-Fr-Fa-0, and is not subjected to excitation. Fr is the force of the forced displacement due to the deflection of the guide rail, and Fa is the force generated by the actuator 11-1. The same applies to the upper active shoe 7-2 when it is subjected to the displacement D2 (t).

Such a process is schematically shown in fig. 10. That is, the feedforward control module 33 generates the feedforward control signals 34-1 and 34-2 … by multiplying and adding the shifts D1(t) and D2(t) … obtained as the rail shift estimation signals 32-1 and 32-2 … and their differential values D1'(t) and D2' (t) … by the spring constant K and the damping constant C, respectively.

As shown in fig. 8, finally, the vibration control signals 38-1, 38-2 …, which are the sums of the feedback control signals 36-1, 36-2 … and the feedforward control signals 34-1, 34-2 …, are supplied to the driving devices 21-1, 21-2 …. Thus, the actuators 11-1 to 11-4 of the active guide shoes 7-1 to 7-4 move in a direction in which horizontal vibration of the car 5 is suppressed.

Next, the deflection (bending) of the guide rails 2-1 and 2-2, which is a basis of the "guide rail displacement model", will be described.

Fig. 11 is a diagram for explaining a feature of a deflected state of the guide rail.

As described above, the "rail displacement model" is modeled assuming that the deflection (bending) of the rails 2-1 and 2-2 changes according to a substantially sinusoidal characteristic having a period corresponding to the installation interval of the carriage 3.

A system for estimating the displacement D1(t) received by the upper active shoe 7-1, the displacement D2(t) received by the lower active shoe, and dD1/dt ═ D1'(t) and dD2/dt ═ D2' (t) which are time derivatives of these are considered.

Here, as shown in fig. 11, the characteristic of the flexure (bending) of the guide rails 2-1 and 2-2 is that the bending component is maximized by the installation interval of the bracket 3. Therefore, the period of the deflection of the guide rails 2-1, 2-2 applied to the car 5 as the exciting force can be assumed to have a frequency ω uniquely determined by the running speed v of the car 5 and the installation interval of the brackets 3₁The characteristic of the periodic sine wave of (a) varies. That is, the following expression is used.

D1(t)＝α·sin(ωt)

Where α is an arbitrary coefficient. In addition, ω is ω₁1/(L/v) × 2 π (rad/s), L being the carrier period [ m [ ]]And v is the running speed [ m/s ]]. The same applies to D2 (t).

The above formula is based on the characteristics of the change in the deflection of the guide rails 2-1, 2-2. That is, as shown in fig. 11, the guide rails 2-1 and 2-2 are normally fixed in the hoistway 1 by the bracket 3 by vertically joining a plurality of guide rail members 2a, 2b, and 2c … each having a predetermined length. Therefore, the possibility that the deflection of the guide rails 2-1, 2-2 varies depending on the arrangement interval of the brackets 3 or the joint of the guide rail members 2a, 2b, 2c … is high.

FIG. 12 is a diagram showing the relationship between the amplitude component and the period of the deflection waveform of the guide rails 2-1 and 2-2.

The deflection waveform of the guide rails 2-1, 2-2 contains a frequency ω determined by the period of the set interval of the carriage 3 and the running speed v₁And a frequency ω determined by the period of the joint of the rail members 2a, 2b, 2c … and the running speed v₂The composition of (1). Wherein the frequency omega₁The most significant component (c) is. Focusing on the frequency omega₁In the case of the component (b), the 2 nd order differential of the shift D1(t) becomes D1 ″ (t) — ω²D1 (t). Here, ω ═ ω₁。

The same applies to the shift D2 (t). When the shift D2(t) is expressed in the form of a state equation, the above equations (6) and (7) are obtained. When equations (6) and (7) are combined with equations (4) and (5) which are state equations of the car vibration model, the extended state equations represented by equations (8) and (9) are formed.

Fig. 13 is a graph showing the comparison between the deflection amount of the guide rail and the deflection amount estimated by the method of the present embodiment, in which the horizontal axis represents time [ sec ] and the vertical axis represents displacement [ mm ].

The waveform 50 shown by the solid line in the figure shows the result of simulating the amount of deflection of the guide rails 2-1, 2-2 that causes vibration of the car 5 during traveling. The waveform 51 indicated by a one-dot chain line in the figure shows the result of simulating the amount of deflection of the guide rails 2-1, 2-2 theoretically estimated by the guide rail displacement estimation module 31. As can be seen from the comparison of the two, the method of the present embodiment can obtain a result similar to the actual deflection amount of the guide rails 2-1 and 2-2.

Fig. 14 is a graph showing the vibration of the car 5 compared with the result of vibration suppression by the method of the present embodiment, in which the horizontal axis represents time [ sec ] and the vertical axis represents acceleration [ gal ].

The waveform 52 shown by the solid line in the figure shows the result of simulating the horizontal vibration generated when the car 5 is forcibly displaced due to the deflection of the guide rails 2-1, 2-2. The waveform 53 indicated by a one-dot chain line in the figure shows the result of simulating the state in which the horizontal vibration is suppressed by the method of the present embodiment. As can be seen from the comparison of the two, the horizontal vibration of the car 5 can be reduced to a state close to 0 by the method of the present embodiment.

As described above, according to the present embodiment, the deflection amount of the guide rails 2-1, 2-2 can be estimated substantially in real time during traveling, and the actuators 11-1 to 11-4 as the vibration damping means can be feedforward controlled. Therefore, even if the deflection of the guide rails 2-1, 2-2 changes with the temperature, humidity, and aging, the horizontal vibration caused by the current deflection can be reliably captured and effectively reduced.

In particular, in the present embodiment, by using a mathematical model that theoretically represents the relative displacement between the guide rail and the car as an observer, the signal of the displacement sensor can be input to the observer to estimate the guide rail displacement (amount of deflection). Therefore, compared to a vibration damping system using an expensive acceleration sensor, there is an advantage that a highly accurate vibration damping system can be realized with a low-cost configuration.

(embodiment 2)

Next, embodiment 2 will be explained.

Fig. 15 is a block diagram showing a functional configuration of the control device 20 according to embodiment 2. Note that the same components as those in fig. 8 of embodiment 1 are denoted by the same reference numerals, and description thereof is omitted.

In embodiment 2, the control device 20 includes two rail displacement estimation modules 31a and 31 b. As shown in fig. 12, when the deflection waveforms of the guide rails 2-1, 2-2 are analyzed, there is a feature that the following two periods are significant.

(1) Cycle of carriage

(2) Rail joint cycle

In embodiment 1 described above, a particularly significant carrier period is modeled as the period of the rail deflection waveform. In this case, since the influence of the rail joint is not taken into consideration, it is also expected that the estimation error increases.

Therefore, in embodiment 2, the following configuration is adopted: two rail displacement estimation modules 31a and 31b are provided, and estimation calculation is performed with attention paid to the bracket period in one rail displacement estimation module 31a and with attention paid to the rail joint period in the other rail displacement estimation module 31 b.

That is, in the rail displacement estimation module 31a, the estimation calculation process is performed using the 1 st rail displacement model assumed to have a characteristic of a substantially sinusoidal wave having the frequency ω determined by the carrier period and the traveling speed v₁. On the other hand, in the rail displacement estimation module 31b, estimation calculation processing is performed using a 2 nd rail displacement model assuming that the rail deflection has a characteristic of a substantially sinusoidal wave having a frequency ω determined by the rail joint period and the running speed v₂。

The car vibration model is similar to that of embodiment 1. That is, the guide rail displacement estimation module 31a estimates the amount of deflection of the guide rails 2-1 and 2-2 based on the relative displacement signals 16-1 to 16-4 output from the displacement sensors 15-1 to 15-4 using the 1 st extended equation of state model in which the car vibration model and the 1 st guide rail displacement model are combined. The rail displacement estimation module 31b estimates the amount of deflection of the rails 2-1, 2-2 based on the relative displacement signals 16-1 to 16-4 output from the displacement sensors 15-1 to 15-4 using a 2 nd extended equation of state model in which a car vibration model and a 2 nd rail displacement model are combined.

In the feedforward control block 33, feedforward control is performed using, as a final estimation result, a result of adding the 1 st rail shift estimation signals 32a-1, 32a-2 … and the 2 nd rail shift estimation signals 32b-1, 32b-2 … output from the rail shift estimation blocks 31a, 31b, respectively.

Similarly, the feedback control module 35 performs feedback control using the result of adding the 1 st rail shift estimation signals 32a-1 and 32a-2 … and the 2 nd rail shift estimation signals 32b-1 and 32b-2 … as the final estimation result.

As described above, according to embodiment 2, by using the rail displacement model in which the two characteristic periods, i.e., the carriage period and the rail joint period are taken into consideration, it is possible to perform the estimation process of the characteristic more reflecting the deflection of the rails 2-1 and 2-2. By performing feedforward control of the actuators 11-1 to 11-4 using the estimation result, a vibration damping effect with higher accuracy can be expected.

In embodiment 2, the feedforward control is performed using the estimation results of both the rail displacement estimation module 31a and the rail displacement estimation module 31b, but the feedforward control may be performed using the estimation result of either one. The same applies to feedback control.

As another method, the guide rail deflection may be modeled focusing on the natural frequency ω n of the horizontal vibration of the car 5. That is, the dominant frequency of the horizontal vibration generated when the car 5 travels at high speed is the resonance frequency ω n [ rad ] possessed by the car 5.

The resonance frequency ω n can be calculated approximately by the following equation (10).

Wherein K is the spring constant [ N/M ] of the upper and lower active guide shoes, and M is the car weight (kg).

When the guide rail deflection is modeled by focusing on such a resonance frequency ω n, the deflection does not match the actual deflection waveform. However, the component of the resonance frequency ω n in the deflection waveform, even if small, causes the car 5 to largely shake. Therefore, there is a possibility that the vibration can be effectively suppressed by modeling at the resonance frequency ω n.

Further, only the component matching the resonance frequency ω n of the car 5 is estimated, and therefore the estimated waveform is small. Therefore, even if the operation amounts of the actuators 11-1 to 11-4 are reduced, the energy saving effect can be expected.

(embodiment 3)

Next, embodiment 3 will be explained.

Fig. 16 is a diagram schematically showing the structure of an active vibration damping device for an elevator according to embodiment 3. Note that the same reference numerals are given to the same parts as those in fig. 1 of embodiment 1, and the description thereof is omitted.

In the above embodiment 1, the displacement sensors 15-1 to 15-4 are provided at 4 positions of the car 5, i.e., the upper, lower, left, and right positions, and the relative displacement is detected at each position. In contrast, in embodiment 3, as shown in fig. 16, displacement sensors 15-1 and 15-4 are provided at two locations, i.e., the upper right side and the lower left side of the car 5. The displacement sensor 15-1 detects the relative displacement between the car 5 and the guide rail 2-1 at the location where the active shoe 7-1 is disposed. The displacement sensor 15-4 detects the relative displacement between the car 5 and the guide rail 2-2 at the location where the active shoe 7-4 is disposed.

The relative displacement signals output from the displacement sensors 15-1, 15-4 are input to the control device 20. The control device 20 uses a mathematical model that theoretically represents the relative displacement as an observer to estimate the amount of deflection of the guide rails 2-1, 2-2 based on the relative displacement signals of the displacement sensors 15-1, 15-4.

In this case, the result of the estimation can be obtained as the rail displacement at the portion where the guide wheel 8-1 of the active shoe 7-1 contacts the one guide rail 2-1 and the rail displacement at the portion where the guide wheel 8-4 of the active shoe 7-4 contacts the other guide rail 2-2. The control device 20 drives the actuators 11-1, 11-4 based on the estimation result. Thus, when the car 5 vibrates in the horizontal direction, the actuators 11-1 to 11-4 work in a direction to suppress the vibration.

As described above, even if the displacement sensors 15-1 and 15-4 are provided at two locations, i.e., the upper right side and the lower left side of the car 5, the deflection amounts of the guide rails 2-1 and 2-2 can be estimated substantially in real time during traveling, and the horizontal vibration of the car 5 can be reduced, as in embodiment 1.

In the example of fig. 16, the displacement sensors 15-1 and 15-4 are provided at two positions, i.e., the upper right side and the lower left side of the car 5, but the displacement sensors 15-2 and 15-3 may be provided at two positions, i.e., the lower right side and the upper left side of the car 5. Alternatively, the displacement sensors 15-1 and 15-3 may be provided at two positions, i.e., the upper right side and the upper left side of the car 5, or the displacement sensors 15-2 and 15-4 may be provided at two positions, i.e., the lower right side and the lower left side of the car 5. In short, when two displacement sensors are used, the displacement sensors may be arranged so as to be able to detect relative displacement between the one guide rail 2-1 side and the other guide rail 2-2 side of the car 5.

(embodiment 4)

Next, embodiment 4 will be described.

Fig. 17 is a diagram schematically showing the structure of an active vibration damping device for an elevator according to embodiment 4. Note that the same reference numerals are given to the same parts as those in fig. 1 of embodiment 1, and the description thereof is omitted.

In embodiment 4, the car 5 is configured by a car frame 6 and a car room 12 enclosing the car frame 6. The car room 12 is a portion on which passengers actually sit, and is connected to the car frame 6 via vibration-proof rubbers 13-1 and 13-2.

Between the car frame 6 and the car room 12, actuators 14-1 and 14-2 for suppressing relative vibration therebetween are interposed.

Further, the active shoes 7-1 to 7-4 are provided with displacement sensors 15-1 to 15-4 as in embodiment 1. Further, in embodiment 4, displacement sensors 15-5 and 15-6 are provided in the upper and lower portions of the car room 12. The displacement sensor 15-5 is provided above the car room 12, and detects a relative displacement in the horizontal direction between the car room 12 and the car frame 6. The displacement sensor 15-6 detects a relative displacement in the horizontal direction between the car room 12 and the car frame 6 at the lower portion of the car room 12.

In fig. 17, L1 is the distance between the center of gravity of the car frame and the upper guide shoe, L2 is the distance between the center of gravity of the car frame and the lower guide shoe, L3 is the distance between the center of gravity of the car frame and the actuator of the upper car frame, L4 is the distance between the center of gravity of the car frame and the actuator of the lower car frame, L5 is the distance between the center of gravity of the car room and the actuator of the upper car frame, and L6 is the distance between the center of gravity of the car room and the actuator of the lower car frame.

Here, in embodiment 1 described above, the two-degree-of-freedom vibration system having the horizontal vibration of the center of gravity and the rotational vibration around the center of gravity of the entire car is modeled, taking into consideration the car frame 6 and the car room 12 constituting the car 5 as a whole. In contrast, in embodiment 4, the car frame 6 and the car room 12 are considered as separate bodies, and a four-degree-of-freedom vibration system having horizontal vibration of the center of gravity of the car frame 6 and rotational vibration around the center of gravity, and horizontal vibration of the center of gravity of the car room 12 and rotational vibration around the center of gravity is modeled.

The extended state equation for a four degree of freedom vibration system is shown below.

Where Xw is the horizontal vibration shift of the center of gravity of the car frame, Xs is the horizontal vibration shift of the center of gravity of the car room, θ w is the rotational vibration angle of the center of gravity of the car frame, and θ s is the rotational vibration angle of the center of gravity of the car room. D1 and D2 are guide rail displacements (deflections) of the upper and lower portions of the car frame, and Fu and Fd are exciting forces (air interference forces) directly applied to the upper and lower portions of the car room. A [16 × 16] is an arbitrary matrix, and specific numerical values are omitted.

Equations (11) and (12) are a set of equations, which are an operation equation and an output equation of an extended state equation model (car vibration + guide rail displacement model) of the four-degree-of-freedom vibration system.

The operating equation of the above equation (11) includes state quantities of D1, D2, Fu, and Fd as disturbance shifts. Here, the state quantities corresponding to the upper and lower portions of one guide rail 2-1 are taken as an example, but actually, the state quantities corresponding to the upper and lower portions of the other guide rail 2-2 may be added.

Now, the relative displacement between the car 5 and the guide rail 2-1 detected by the displacement sensor 15-1 is set to h1, and the relative displacement between the car 5 and the guide rail 2-1 detected by the displacement sensor 15-2 is set to h 2. The relative displacement between the car frame 6 and the car room 12 detected by the displacement sensor 15-5 is hu, and the relative displacement between the car frame 6 and the car room 12 detected by the displacement sensor 15-6 is hd.

The output equation of the above equation (12) outputs 8 state quantities (h1', h1, h2', h2, hu ', hu, hd', hd) composed of the above relative displacements and the relative displacement speed that is the 1 st order differential thereof. Similarly to fig. 6, when these state quantities are compared with the actual displacement sensor signal and its differential signal, and the difference is multiplied by the observer gain and returned, the rail displacement (D1, D2, Fu, Fd) to be estimated and the time differentials (D1', D2', Fu ', Fd') thereof are obtained. In this case, the displacements D1 and D2 caused by the deflection of the guide rails can be estimated including the exciting forces Fu and Fd exerted on the car chamber 12 mainly by the wind pressure during traveling.

Further, as described in FIG. 7, the signals of the shift sensors 15-1 to 15-6 may be input after being multiplied by a predetermined gain.

Here, for the sake of simplicity of explanation, the case of estimating the amount of deflection of one guide rail 2-1 is assumed and described, but actually, the amount of deflection of the other guide rail 2-2 is also estimated using the displacement sensors 15-3 and 15-4.

Specifically, an observer having an extended state equation of the four-degree-of-freedom vibration system of the above equations (11) and (12) is incorporated in the rail displacement estimation module 31 shown in fig. 8. Then, the signals of the displacement sensors 15-1 to 15-6 are input to the rail displacement estimation module 31, and the displacements D1 and D2 due to the deflection of the rails and the excitation forces Fu and Fd applied to the car room 12 are estimated.

The feedforward control module 33 feedforward-controls the actuators 11-1 to 11-4 and the actuators 14-1 and 14-2 provided between the car frame 6 and the car room 12 based on the estimation result. Thus, for example, when the car room 12 vibrates due to wind pressure or the like, such as when 2 elevators are staggered during high-speed traveling, the actuators 14-1 and 14-2 can be driven in a direction to suppress the vibration, thereby stabilizing the sway of the car room 12.

As described above, according to embodiment 4, by modeling the vibration system as a separate body from the car frame 6 and the car room 12, it is possible to suppress vibration of the car frame 6, which is received as a forced displacement force due to deflection of the guide rails 2-1 and 2-2, and also to suppress vibration due to an excitation force received by the car room 12 during high-speed traveling, by using the displacement sensors 15-1 to 15-6.

(embodiment 5)

Next, embodiment 5 will be described.

In the above embodiment 1, the extended state equation using the equations (8) and (9) constitutes an observer. In this observer, a conversion equation (part (c) of the equation) for relative displacement is required to obtain matching between the actual value of the displacement sensor signal and the output value of the extended state equation. In contrast, in embodiment 5, the observer is configured using a state equation in which a part of the conversion equation is calculated in advance (see fig. 18).

That is, first, a relational expression of the relative shift and the absolute shift shown in expression (13) is used.

X+L1e-D1＝h1

X-L2θ-D2＝h2...(13)

The formula (13) is the same as the formula (3). h1 is the relative displacement between the car 5 and the guide rail 2-1 detected by the displacement sensor 15-1, and h2 is the relative displacement between the car 5 and the guide rail 2-1 detected by the displacement sensor 15-2. D1 represents the displacement (amount of deflection) of the installation location of the upper active shoe 7-1 of the guide rail 2-1, and D2 represents the displacement (amount of deflection) of the installation location of the lower active shoe 7-2 of the guide rail 2-1. (X + L1 theta), (X-L2 theta) represent absolute shifts.

Using such a relational expression, the state quantities (X ', X, θ', θ) of the car vibration obtained from the signals of the acceleration sensors are transformed, and a "car displacement model" indicating the state quantities (h1', h2', h1, h2) of the car displacement is prepared. In embodiment 5, an "extended equation of state model" is used, which is obtained by combining the "car shift model" with the "guide rail shift model". The "extended equation of state model" is expressed by the following equations (14) and (15).

Equations (14) and (15) are a set of equations, which are an operation equation and an output equation of an extended state equation model (car shift + guide rail shift model) of the two-degree-of-freedom vibration system. Further, h1 and h2 are relative displacement amounts (m) between the upper and lower guide rails of the car, and D1 and D2 are guide rail displacement amounts (deflection amounts) (m) between the upper and lower guide rails of the car. U1, U2 are the forces of the actuators of the upper and lower car. AA [ 8X 8], BB [ 8X 2], and CC [ 4X 8] include components of a guide rail displacement model. Since the calculation formula is complicated, the details are omitted.

Here, the difference from the extended equation of state (see equations (8) and (9)) in embodiment 1 is that an equation of state representing the vibration characteristics in the horizontal direction that the car 5 receives due to the deflection of the guide rail 2-1 is modified, and the relative displacement between the guide rail 2-1 and the car 5 is expressed in the form of an equation of state with the state vector as the state vector. When an observer is configured using such an extended state equation, the observer becomes as shown in fig. 18.

Fig. 18 is a diagram for explaining the structure of the scope according to embodiment 5. In the figure, reference numeral 40 shows a real object (an elevator car) representing an actual vibration system. 61 is a viewer.

Now, a case of estimating the displacement of the guide rail due to the deflection of one guide rail 2-1 will be described as an example. During the running of the car 5, signals indicating the relative displacement between the guide rail 2-1 and the car 5 are input from the displacement sensors 15-1, 15-2 to the observer 61.

The observer 61 is composed of a car displacement + guide rail displacement model 62, a difference calculation unit 63, and an observer gain matrix 64.

The car displacement + guide rail displacement model 62 corresponds to the extended state equation shown in the above equations (14) and (15), and outputs the state quantities of car displacement (h1', h2', h1, h2) and the estimated quantities of guide rail displacement (D1', D2', D1, D2). The state quantities of car displacement (h1', h2', h1, h2) output from the car displacement + guide rail displacement model 62 are supplied to the difference calculation unit 63 as estimated values of relative displacement.

The difference calculation unit 63 compares the estimated value of the relative shift with the actually measured value of the relative shift, and feeds back the difference value to the car shift + guide rail shift model 62 via the observer gain matrix 64. The observer gain matrix 64 is a matrix for multiplying a difference value between an estimated value and an actually measured value of the relative shift by a predetermined gain, similarly to the observer gain matrix 45 of fig. 6.

Here, the observer 41 of fig. 6 used in the above embodiment 1 is configured to feed back the estimated amounts (D1', D2', D1, D2) of the rail displacement as output values. That is, since the estimation is performed by feeding back the estimation result, for example, when the estimation result includes an error, the accuracy may be lowered.

On the other hand, the observer 61 according to embodiment 5 has an advantage that estimation accuracy and estimation stability are easily obtained because the estimated amounts (D1', D2', D1, D2) of the rail displacement are not fed back. However, a complicated formula corresponding to the relative shift transformation matrix 43 of fig. 6 needs to be solved in advance.

When the observer 61 having such a structure is incorporated in the control device 20 and signals from the displacement sensors 15-1 and 15-2 are input to the control device 20 as shown in fig. 3, the amount of deflection of the guide rail 2-1 can be estimated. When the actuator 11-1 of the active shoe 7-1 and the actuator 11-2 of the active shoe 7-2 are driven by feedforward control based on the estimation result, displacement due to deflection of the guide rail 2-1 can be absorbed in real time, and horizontal vibration of the car 5 can be reduced.

In the configuration of fig. 8, the signals of the shift sensors 15-1 and 15-2 are input to the observer 41 in the original size, but the signals of the shift sensors 15-1 and 15-2 may be input after being multiplied by the correction gain 46, as shown in fig. 7, for example.

Here, for the sake of simplifying the description, the case of estimating the amount of deflection of one guide rail 2-1 is assumed and described, but in practice, the amount of deflection of the other guide rail 2-2 is also estimated using the displacement sensors 15-3 and 15-4.

In short, the observer 61 estimates the amount of deflection of the guide rails 2-1 and 2-2 in substantially real time using an extended equation of state model, using the relative displacement between the guide rails and the car obtained as the state quantity of the car 5 and the relative displacement speed as input signals, based on the signals of the displacement sensors 15-1 to 15-4. By performing feedforward control of the vibration damping mechanisms (actuators 11-1 to 11-4) of the active shoes 7-1 to 7-4 based on the estimation result of the observer 41, it is possible to obtain a vibration damping effect similar to the method in which the deflection amounts of the guide rails 2-1 and 2-2 are learned in advance.

Here, for the sake of simplifying the description, the case of estimating the amount of deflection of one guide rail 2-1 is assumed and described, but in practice, the amount of deflection of the other guide rail 2-2 is estimated including the amount of deflection of the other guide rail using the displacement sensors 15-3 and 15-4.

In short, the observer 41 estimates the amount of deflection of the guide rails 2-1 and 2-2 in substantially real time using an extended equation of state model, using the relative displacement and the relative displacement speed between the guide rails and the car, which are obtained as the state quantities of the car 5, as input signals based on the signals of the displacement sensors 15-1 to 15-4. By performing feedforward control of the vibration damping mechanisms (actuators 11-1 to 11-4) of the active shoes 7-1 to 7-4 based on the estimation result of the observer 41, it is possible to obtain a vibration damping effect similar to the method in which the deflection amounts of the guide rails 2-1 and 2-2 are learned in advance.

As described above, according to embodiment 5, the observer is configured by using the extended state equation obtained by combining the car shift model, which has the relative shift between the guide rail and the car as the state vector, and the guide rail shift (the amount of deflection) as the state vector, and the guide rail shift model, which has the relative shift between the guide rail and the car as the state vector, to estimate the guide rail shift with higher accuracy and improve the vibration damping effect.

In addition, this embodiment 5 and the above-described embodiments 2 to 4 can also be combined as appropriate.

According to at least one embodiment described above, it is possible to provide an active shoe device capable of reliably capturing and effectively reducing vibration caused by deflection of a guide rail during car travel with an inexpensive configuration.

In the above embodiments, the elevator is described as an example, but the method of the present invention can be applied to reduce vibration as long as the moving object travels on a guide rail, such as an electric train.

In short, although several embodiments of the present invention have been described, these embodiments are presented as examples and are not intended to limit the scope of the invention. These new embodiments can be implemented in various other ways, and various omissions, substitutions, and changes can be made without departing from the spirit of the invention. These embodiments and modifications thereof are included in the scope and gist of the invention, and are included in the invention described in the claims and the scope equivalent thereto.

Claims

1. An active vibration damping device for an elevator, comprising:

a cage that moves up and down along the guide rail;

2. Active vibration damping device of an elevator according to claim 1,

the vibration damping mechanism is provided on at least one of a portion of the upper part of the car facing the guide rail and a portion of the lower part of the car facing the guide rail,

the displacement sensor detects a relative displacement between the car and the guide rail at a position where the vibration damping mechanism is provided.

3. Active vibration damping device of an elevator according to claim 1,

the mathematical model includes an extended equation of state model obtained by combining a car vibration model, which is a model expressing the vibration characteristics of the car in the horizontal direction due to the deflection of the guide rail, and a guide rail displacement model, which is a model expressing the deflection of the guide rail in the form of an equation of state on the assumption that the deflection of the guide rail changes in a predetermined regular characteristic,

the extended equation of state model includes a conversion expression for calculating a relative displacement between the guide rail and the car based on a relationship between a state quantity of the car and an estimated quantity of deflection of the guide rail.

4. Active vibration damping device of an elevator according to claim 1,

the mathematical model includes an extended equation of state model obtained by combining a car displacement model in which an equation of state representing the vibration characteristics in the horizontal direction to which the car is subjected by the deflection of the guide rail is transformed and a guide rail displacement model in which the deflection of the guide rail is assumed to change with a predetermined regular characteristic and which is expressed in the form of an equation of state, and the relative displacement between the guide rail and the car is expressed in the form of a state vector.

5. Active vibration damping device of an elevator according to claim 3 or 4,

the estimating means is constituted by an observer that compares the relative displacement between the car and the guide rail detected by the displacement sensor with a theoretically calculated relative displacement obtained as an output value of the extended state equation model, and feeds back a difference between the relative displacement and the theoretically calculated relative displacement.

6. Active vibration damping device of an elevator according to claim 5,

the observer multiplies and inputs a signal indicating the relative displacement between the car and the guide rail detected by the displacement sensor by a predetermined gain.

7. Active vibration damping device of an elevator according to claim 3 or 4,

the guide rail displacement model is modeled on the assumption that the deflection of the guide rail has a substantially sinusoidal characteristic in which only the amplitude and the phase change at a constant cycle.

8. Active vibration damping device of an elevator according to claim 7,

the guide rail displacement model is modeled assuming that the deflection of the guide rail has a characteristic of a substantially sinusoidal wave having a period at an installation interval of brackets for fixing the guide rail in the hoistway.

9. Active vibration damping device of an elevator according to claim 7,

the guide rail displacement model is modeled assuming that the deflection of the guide rail has a characteristic of a substantially sinusoidal wave having a period at intervals of joints of a plurality of guide rail members constituting the guide rail.

10. Active vibration damping device of an elevator according to claim 7,

the guide rail displacement model is modeled assuming that the deflection of the guide rail has a characteristic of a substantially sinusoidal wave having a 1 st cycle at an installation interval of brackets for fixing the guide rail in the hoistway and a 2 nd cycle at an interval of joints of a plurality of guide rail members constituting the guide rail.

11. Active vibration damping device of an elevator according to claim 3,

the car vibration model is a model obtained by modeling a vibration system having two degrees of freedom of vibration, horizontal vibration of the center of gravity of the car and rotational vibration around the center of gravity.

12. Active vibration damping device of an elevator according to claim 3,

the car vibration model is a model obtained by modeling a vibration system having 4 degrees of freedom, i.e., horizontal vibration of the center of gravity of the car frame, rotational vibration around the center of gravity, horizontal vibration of the center of gravity of the car chamber, and rotational vibration around the center of gravity, while the car frame and the car chamber are formed as separate bodies.

13. Active vibration damping device of an elevator according to claim 12,

the estimating means estimates a vibration force applied to the car chamber using the car vibration model,

the control means performs feed-forward control of an actuator provided between the car frame and the car room in a direction in which an excitation force applied to the car room is suppressed, based on a result of the estimation by the estimation means.