JPH0743252A - Load test equipment - Google Patents

Abstract

PURPOSE:To allow highly accurate constant speed test and constant power test corresponding to the load conditions under actual operation of a sample rotating machine. CONSTITUTION:The load test equipment comprises a hydraulic brake, a hydraulic pressure regulating means, a servo amplifier for controlling the pressure regulating means, and a torque command modifying section 6 disposed on the input side of a transmission element 8 for converting a torque command signal Es into a brake torque Tb while comprising a PI operating unit 60 receiving the difference between the rotational speed signal En and a torque command signal Es of a sample rotating machine, an analog memory 64 storing a reference target torque, and an adder 65 for determining the difference between both outputs. An output from the torque command modifying section 6, i.e., an output from the adder 65, is delivered to the servo amplifier and load test is performed under a constant speed. The load test equipment further comprises a PI operating unit 61 receiving the product of the torque command Es and a brake torque Tb from the brake and the adder 65 delivers the difference signal between the output from the PI operating unit 61 and a value stored in the analog memory 64 to the servo amplifier thus performing load test under constant power.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a load testing apparatus used for load testing of rotating machines. More specifically, a hydraulically actuated brake is used as a means for generating a load torque applied to a test rotating machine. Load test apparatus.

[0002]

2. Description of the Related Art A load testing apparatus using a hydraulically actuated brake as a load torque generating means is disclosed, for example, in Japanese Utility Model Application Laid-Open No. 2-21538 and Japanese Patent Application Laid-Open No. 3-56837. The brake is a so-called wet multi-plate brake, which is provided with a housing and a rotation shaft rotatably supported inside the housing, and has a large number of coaxially mounted rotation-restricted outsides of the rotation shaft. Of the rotary braking plate, and a large number of fixed braking plates similarly mounted inside the housing,
Alternately polymerized through the oil enclosed in the housing,
It is configured such that these are pressed against each other by an operating cylinder operated by hydraulic pressure to generate a braking torque.

In the load test by the load testing device, the housing of the brake constructed as described above is supported via a swing arm of a predetermined length which is provided on the outer side of the brake housing, and the rotary shaft is also tested. The load applied to the support portion of the swing arm (load equivalent to the braking torque) by the reaction of the braking torque generated by the brake is detected by the interlocking connection with the machine, and the feedback signal of the detection result and the torque command given from the outside are detected. The brake hydraulic pressure supplied to the working cylinder is regulated based on the signal to adjust the braking torque generated by pressing the working cylinder, and the desired load torque corresponding to the torque command signal is supplied to the working cylinder. It is performed with a load on the trial rotating machine.

The load testing devices disclosed in Japanese Utility Model Laid-Open No. 2-21538 and Japanese Patent Laid-Open No. 3-56837 are interposed in the supporting portion of the swing arm to apply the braking torque equivalent load in one direction. A hydraulic control valve having a spool, and a current control type pressure control valve for generating a control hydraulic pressure for urging the spool in the opposite direction are provided, and a configuration for adjusting the brake hydraulic pressure by these is provided. Has become. The feedback signal of the load equivalent to the braking torque is given to the servo amplifier which performs the control operation of the pressure control valve together with the torque command signal, and the servo amplifier gives a deviation between the two signals, that is,
An operation of controlling the operating current of the pressure control valve is performed in order to eliminate the deviation between the braking torque actually generated by the brake and the target torque.

That is, the spool of the hydraulic control valve has a load equivalent to the braking torque acting from one side through the swing arm and a pressing load acting on the other end surface by the control hydraulic pressure sent from the pressure control valve. Maintaining a balanced position, the brake will generate a braking torque corresponding to the torque command signal. Further, in this state, when the braking torque generated by the brake changes due to various disturbances, the spool is displaced in the acting direction of the braking torque equivalent load,
At this time, the control operation of the servo amplifier according to the deviation generated between the feedback signal of the braking torque and the torque command signal increases or decreases the control oil pressure sent from the pressure control valve. , The brake is quickly returned to the equilibrium position before the displacement, and the brake constantly continues to generate the braking torque.

On the other hand, when the torque command signal is changed, the control oil pressure is increased or decreased so as to eliminate the deviation between the changed torque command signal and the current braking torque. When the brake is displaced to the equilibrium position, the brake generates a braking torque corresponding to the changed torque command signal and loads the test rotating machine. in this way,
By appropriately changing the torque command signal, the load test under constant torque of the test rotating machine is performed.

Further, in the load testing apparatus disclosed in Japanese Patent Laid-Open No. 3-56837, a rotation speed detector for detecting the rotation speed of the test rotating machine that rotates with the brake is attached to a part of the rotation axis of the brake. At the same time, the torque command signal is continuously changed according to the change in the rotation speed of the test rotating machine detected by the rotation speed detector and the analog memory holding the torque command signal on the input side of the servo amplifier. It is equipped with a torque command change unit, and is mainly used for the load test under constant torque described above.The constant speed test is performed with the rotation speed of the test rotating machine kept constant, and the absorption power by the brake is kept constant. The load test can be performed by simulating various load states during actual operation of the rotating machine under test, such as a constant power test performed in such a state.

In the constant speed test and the constant power test, an appropriate torque command signal is given to the servo amplifier to realize a constant torque load state, and then constant speed control or constant power control is selected by a predetermined switch operation. It is done by doing.
When these selections are made, the analog memory holds the current torque command signal, and the servo amplifier is given a torque command signal obtained by correcting the value held in the analog memory by the output of the torque command changing unit. It is like this.

[0009]

However, the torque command changing unit disclosed in Japanese Patent Laid-Open No. 3-56837 is an imperfect one that does not correspond to the configuration of the entire control system.
When the constant speed test or constant power test is performed by the control operation of the servo amplifier based on the output of the torque command changing unit, since an approximate first-order delay in the hydraulic piping system leading to the working cylinder of the brake is not considered, There is a practical problem that the response of the braking torque generated by the brake is low with respect to the change of the control signal which is the output of the servo amplifier, and the reliability of the obtained test result is low.

The present invention has been made in view of such circumstances, and a load capable of carrying out a constant speed test and a constant power test with high accuracy in order to correspond to a load state during actual operation of a test rotating machine. The purpose is to provide a test device.

[0011]

A load test apparatus according to a first aspect of the present invention is a hydraulically operated brake interlockingly connected to a rotating machine under test, and a means for adjusting a brake hydraulic pressure supplied to the brake. And a servo amplifier that controls the pressure adjusting means based on a feedback signal of a braking torque generated by the brake and a torque command signal given from the outside,
A torque command changing unit that adjusts the torque command signal according to a change in the rotation speed of the test rotating machine, and by giving the output of the torque command changing unit to the servo amplifier, In a load test apparatus capable of performing a load test at a constant speed, the torque command changing unit serves as a reference and a PI calculator that receives a deviation between the torque command signal and the rotation speed detection signal. It is characterized by comprising an analog memory for holding an average target torque and an adder for obtaining a deviation between the outputs of the both and giving it to the servo amplifier.

The load testing apparatus according to the second aspect of the present invention is a hydraulically operated brake interlockingly connected to the rotating machine under test, a means for adjusting the brake hydraulic pressure supplied to the brake, and an external device. A servo amplifier that controls the pressure adjusting means based on a given torque command signal and a feedback signal of the braking torque generated by the brake, a feedback signal of the braking torque, and a detection signal of the rotation speed of the test rotating machine. And a torque command changing unit that adjusts the torque command signal according to a change in the absorbed power due to the brake, and the output of the torque command changing unit is given to the servo amplifier, thereby providing the servo amplifier. In a load testing device capable of performing a load test of a test rotating machine under constant power, the torque command changing unit includes the torque command signal and the multiplication command. It is provided with a PI calculator for inputting a deviation from a value, an analog memory for holding a reference average target torque, and an adder for calculating a deviation between outputs of both and outputting it to the servo amplifier. And

[0013]

In the present invention, the control amount in the constant speed control is determined based on the PI calculation value which receives the deviation between the rotation speed of the DUT and the torque command signal, and the control in the constant power control is performed. According to a configuration in which the amount is determined based on a PI calculation value that takes a deviation between a product of the rotation speed of the test rotor and the braking torque generated by the brake (absorption power by braking) and the torque command signal as an input. The transfer function of the entire control system in the constant speed control and the constant power control has a form similar to that in the constant torque control, that is, a quadratic delay system to improve the responsiveness.

[0014]

DESCRIPTION OF THE PREFERRED EMBODIMENTS The present invention will be described below in detail with reference to the drawings showing the embodiments. FIG. 1 is a schematic diagram showing the configuration of a hydraulic circuit of a load test apparatus according to the present invention. As shown in the figure, the load test apparatus is a hydraulically operated brake 1 that generates a load torque, and braking and release of the brake 1. to occupy perform the operation, the hydraulic control valve 2 for controlling the feed oil pressure to the actuation cylinder 10 arrangement (the brake pressure P _b) in the interior of this, the hydraulic pump 30 is a source of brake fluid pressure P _b, the hydraulic control A hydraulic pump 31 that is a source of a control hydraulic pressure P _c that is added to the spool 20 of the valve 2 as described below, and is interposed between the hydraulic pump 31 and the hydraulic control valve 2 to distribute the control hydraulic pressure P _c . And a current control type pressure control valve 32 for controlling

The brake 1 supports a rotary shaft 11 which is interlocked with the output shaft of the test rotary machine A, and supports the rotary shaft 11.
A wet multi-disc brake including a housing (12) in which oil is enclosed, and a braking torque is generated by a shear resistance of an oil film between a large number of braking discs attached to both of them while restraining their rotation. 2 is a vertical cross-sectional view showing the internal structure of the brake 1, and FIG. 3 is an enlarged view of a main part of FIG. The upper half of FIG. 2 shows the braking state, and the lower half of FIG. 2 and FIG. 3 show the braking released state.

The housing 12 of the brake 1 is swingably supported between a pair of bearings C and C, which are vertically erected on a base (not shown) with a proper distance therebetween, through separate bearings 12A and 12B. Has been done. The rotating shaft 11 of the brake 1 is a housing
12 is inserted from one side (the left side in the figure), and the insertion side, that is, the middle part of the left side is fitted and fixed to the housing 12 by the bearings 11A and 11B. Each of them is supported by a bearing 11C that is fitted and fixed, and can rotate about its axis.

The housing 12 is provided with an inner housing 14 which surrounds the outer side of the rotary shaft 11.
A large number of braking plates are attached to the inner circumference of the so as to be restrained from rotating and movable in the axial direction. In addition, in the surrounding portion of the inner housing 14 on the outer periphery of the rotating shaft 11,
Rotation is restricted by the rotary shaft 11, and a large number of braking plates are attached so as to be movable in the axial direction.
The braking plates on the inner circumference of the inner housing 14 are alternately superposed in the axial direction.

At the right end of the housing 12, the bearing
A rotary joint 15 is fitted through 15A, and the rotary joint 15 and an oil passage continuous with the rotary joint 15 are provided inside the inner housing 14.
Oil introduced and discharged via 16 is enclosed. The left end of this oil passage 16 is an oil passage that penetrates the center of the rotating shaft 11.
16a, the bearings 11A and 11B are communicated with the positions where the bearings 11A and 11B are arranged. Has become.

In the brake 1, when the braking plates on the rotary shaft 11 side and the inner housing 14 side are brought close to each other, the braking torque is generated by the shear resistance of the oil film generated between the two through the enclosed oil inside the inner housing 14. When they are generated and are separated from each other, the braking torque is released.
On the right side of the inner housing 14, a double-acting working cylinder 10 for pressing a brake plate to perform the approach is formed. The working cylinder 10 is provided with two oil chambers 10a and 10b, and a piston 100 that receives internal pressures of these oil chambers on both sides thereof.

As shown in FIG. 3, the working cylinder 10 has a stepped ring hole (a larger diameter on the braking plate side and a smaller diameter on the farther side) provided around the right end portion of the inner housing 14,
A corresponding stepped annular piston 100 is fitted. The oil chamber 10a on one side of the piston 100 is formed on the right side of the large diameter portion of the stepped annular hole, and the oil chamber 10b on the other side is formed on the left side of the small diameter portion, and separate oil passages 17a and 17b are formed. Through the right outer wall of the housing 12 at different positions, and these oil chambers 10a and 10b are connected to the open ends of the oil passages 17a and 17b through separate pipe joints 19a and 19b, respectively. Oil pipe
The brake oil pressure _Pb is introduced through 18a and 18b.

Thus, the piston 100 has one oil chamber 1
Advanced operative to press the brake plate when the brake hydraulic pressure P _b is introduced into 0a, brake hydraulic pressure P to the other oil chamber 10b
_{When b} is introduced, it will move in and out. Disc springs 101, 101, ... for urging the brake plates against each other in a separating direction are interposed between the brake plates, and when the piston 100 retracts, each brake plate has a disc spring between them. 101,1
It is separated by the spring force of 01 ..., and the braking is reliably released.

A swing arm 13 of a predetermined length (see FIG. 1) is provided at the center of the housing 12 of the brake 1 constructed as described above so as to project outward in the radial direction. The front end of the hydraulic pressure control valve 2 is supported by the spool 20 of the hydraulic control valve 2 which is vertically fixed on the base. spool
A load cell 41 for detecting the equivalent load of the braking torque applied to the spool 20 via the swing arm 13 in association with the generation of the braking torque by the brake 1 is provided at the connecting portion between the 20 and the swing arm 13. ing.

The hydraulic control valve 2 as shown in detail in FIG.
It has a built-in cylindrical spool 20 having four large-diameter portions arranged at predetermined intervals in the axial direction, and forms a first oil chamber below the lowermost large-diameter portion. Second, third, and fourth oil chambers are formed between the large diameter portions in order from the bottom, and further a fifth oil chamber is formed above the uppermost large diameter portion. Hydraulic braking pressure P _b of the hydraulic pump 30 is generated is supplied to the third fluid chamber via the pump port 23, also hydraulic said hydraulic pump 31 is generated, a control port 32A and the hydraulic control valve of the pressure control valve 32 2
Is supplied to the first oil chamber via the control port 21 of the control port 32 of the pressure control valve 32 and the control port of the hydraulic control valve 2
It is supplied to the fifth oil chamber via 25. Also, the hydraulic control valve 2
The second oil chamber and the fourth oil chamber are opened to an oil tank T maintained at a low pressure via separate return ports 22 and 24.

Return springs 2A and 2B for urging the spool 20 toward the center are arranged in the first oil chamber located at the bottom and the fifth oil chamber located at the top, respectively. There is no pressure difference in the control oil pressure P _c supplied to the oil chamber via the pressure control valve 32, and the swing arm 13 is provided in order to prevent the swing of the housing 12 caused by the reaction of the braking operation of the brake 1. When there is no external force (braking torque equivalent load) acting as described below, the spool 20 forms a spring center at a position (neutral position) determined by the spring load of the return springs 2A and 2B, and stops. It has become.

The hydraulic control valve 2 has a pair of brake ports 26R, 2 each having an opening at a position where it is closed by the large diameter portions on both sides of the third oil chamber when the spool 20 is in the neutral position.
6F is formed. Thus, spool from the neutral position
When 20 moves downward (or upward), the brake port 26R (or 26F) located on the lower side (or upper side) is the third
The brake oil pressure P _b supplied to the third oil chamber via the pump port 23 is 4 ports 3
It is fed to one of the oil chambers 10a and 10b of the working cylinder 10 through the position switching type electromagnetic switching valve 33 and the separate oil feeding pipes 18a and 18b, and the brake 1 performs braking or braking releasing operation.

On the other hand, at this time, the other brake port 26F (or 26R) of the hydraulic control valve 2 opens in the fourth oil chamber (or the second oil chamber), and the return port 24 (or the return port 22) is used. Open to the oil tank T. Therefore, the hydraulic oil discharged from the other oil chamber 10b, 10a by the above-described operation of the brake 1 returns to the oil tank T via the respective oil feed pipes 18b, 18a, the electromagnetic switching valve 33 and the hydraulic control valve 2, Without damaging the operation of the piston 100 for braking or releasing braking, the braking and releasing operations occur at high speeds. Also, the hydraulic control valve 2
The electromagnetic switching valve 33 between the brake 1 and the brake 1 switches the feed path of the brake oil pressure P _b according to the forward / reverse rotation direction of the rotary shaft 11 connected to the output shaft of the test rotary machine A. It is provided to prevent the occurrence of brake lock.

The spool of the hydraulic control valve 2 having the above structure
When acting on 20 via the swinging arm 13 corresponding to the braking torque corresponding to the braking operation of the brake 1, the return spring 2
The spring load of A and 2B and the first and the fifth through the pressure control valve 32.
Control load F _c corresponding to control oil pressure P _c introduced into the oil chamber
(Refer to FIG. 4) The brake 1 moves in the vertical direction according to the balance with the force, and the brake 1 performs a braking operation or a brake releasing operation by the brake oil pressure P _b supplied according to the movement of the spool 20.

A speed detector 42 is interposed between the cylinder block of the hydraulic control valve 2 and the spool 20. The speed detector 42 is a linear speed converter that generates a speed signal corresponding to the moving speed of the spool 20, and the detection result of the speed detector 42 is the detection result of the load equivalent to the braking torque of the brake 1 by the load cell 41. It is also used as a feedback signal for damping correction of the target torque.

Further, as shown in FIG. 2, a rotation speed detector 43 is fixed on the support C located on the protruding side of the rotary shaft 11 in the brake 1, and the rotary shaft 11 is projected through a timing belt 44. Connected to the department. As the rotation speed detector 43, the rotating shaft 11 and the test rotating machine A that rotates together with the rotating shaft 11 are used.
A rotary encoder that emits a predetermined number of pulse signals per revolution is used. This pulse signal is converted into a voltage signal corresponding to the rotation speed of the test rotating machine A as described later, and constant speed and constant power control is performed. Is used as a feedback signal for changing the torque command signal.

In the load test by the load test apparatus having the above-mentioned structure, the output shaft of the test rotary machine A is set to the rotary shaft 11 of the brake 1.
After the preparatory work for changing the switching position of the electromagnetic switching valve 33 according to the rotation direction of the test rotating machine, a control signal corresponding to a desired target torque is given to the drive circuit of the pressure control valve 32. This is performed by changing the operating current to the pressure control valve 32. As a result, the hydraulic pressure generated by the hydraulic pump 31 is introduced into the first and fifth oil chambers of the hydraulic control valve 2 as the control hydraulic pressure P _c having a distribution ratio determined by the operation of the pressure control valve 32. Then, a pressure difference corresponding to the target torque occurs.

For example, when the pressure difference of the control oil pressure P _c is generated from the first oil chamber to the fifth oil chamber and the electromagnetic switching valve 33 is switched to the lower position, the spool 20 changes to the control oil pressure P _c . by the action of the corresponding control force F _c, back moves upward from the neutral position against the bias of spring 2B, the brake port 26F is the third oil chamber, the brake port 26R is respectively open to the second oil chamber To do. As a result, from the hydraulic pump 30 to the third
Hydraulic braking pressure P _b supplied to the oil chamber, the brake port
The oil is fed to the advancing side oil chamber 10a of the working cylinder 10 through the 26F, the electromagnetic switching valve 33 and the oil feeding pipe 18a, while the retreating side oil chamber 1
The hydraulic oil inside 0b is returned to the second oil chamber through the oil feed pipe 18b, the electromagnetic switching valve 33 and the brake port 26R, and is discharged to the oil tank T. As a result, the working piston 100 rapidly advances, the braking plates on the rotating shaft 11 side and the inner housing 14 side are pressed against each other, and the brake 1 generates braking torque due to the shear resistance of the oil film between the braking plates. It is assumed that the housing 12 is not rotated in the rotation direction of the rotating shaft 11.

At this time, the braking torque generated by the brake 1 is converted into a braking torque equivalent load through the swing arm 13 and applied to the spool 20.
Is a balance between the braking torque equivalent load acting downward, the control load F _c acting upward due to the control oil pressure P _c introduced into the first and fifth oil chambers, and the spring loads of the return springs 2A and 2B. Stop in position. In this state, when the braking torque generated by the brake 1 fluctuates due to disturbance, the spool 20 of the hydraulic control valve 2 also moves up and down, which changes the opening mode of the brake ports 26R, 26F and the piston 100 of the working cylinder 10 moves. As a result of the inching, the braking torque generated by the brake 1 automatically returns to the torque before the change, that is, the target torque.

Such movement of the spool 20 occurs even when the setting of the target torque is changed and the control oil pressure P _c introduced into the first and fifth oil chambers of the oil pressure control valve 2 is changed.
In this case as well, the same operation changes the braking torque of the brake 1, and at the time when this changes to the changed target torque, the spool 20 reaches a new equilibrium position and the brake 1
Generate a braking torque corresponding to the equilibrium position, that is, a braking torque corresponding to the newly set target torque.

The operation of the load test apparatus according to the present invention has been described above by paying attention to the oil flow, but in this apparatus, the load equivalent to the braking torque generated by the brake 1 is caused by the load cell 41, and this load fluctuates. The moving speed of the spool 20 of the hydraulic control valve 2 added via the arm 13 is detected by the speed detector 42, respectively, and the feedback control based on the detection result is performed. Further, the rotation speed of the test rotating machine A that rotates while being applied with the braking torque by the brake 1 is detected by the rotation speed detector 43, and the target torque is changed using this detection signal to determine the constant speed and the constant power. The load test below is performed.

FIG. 4 is a block diagram of the control system of the load testing apparatus according to the present invention. In the constant torque control which is a basic control function, the target torque is given to the servo amplifier 5 as a voltage signal (torque command signal E _s ) corresponding to this magnitude.

The servo amplifier 5 also has a torque feedback signal E _b detected by the load cell 41 and corresponding to the braking torque equivalent load, and a speed feedback signal E _b detected by the speed detector 42 and corresponding to the moving speed of the spool 20. ₀ is given respectively, and the servo amplifier 5
Processes these signals as described later and outputs a control signal to the pressure control valve 32.

The specific operation of the pressure control valve 32 is performed according to the control signal electrically given from the servo amplifier 5,
The throttle opening corresponding to each of the pair of control ports 32A, 32B is changed. When this change is made, a difference occurs in the control hydraulic pressure P _c on the downstream side of both control ports 32A, 32B.
This control oil pressure P _c acts on the spool 20 of the oil pressure control valve 2. That is, the control signal output from the servo amplifier 5 includes the pressure control valve 32 and the hydraulic piping system connected to the pressure control valve 32, and as shown in FIG. 4, passes through the charge-weight conversion unit 34 represented as a proportional element having _a proportional sensitivity K _a. Converted to the dimension of force, control load F
It becomes _c and acts on the spool 20.

The hydraulic control valve 2 is replaced with a mechanical addition point 27, an inertial delay element 28 and a pressure conversion section 29. The mechanical addition point 27 corresponds to the spool 20. At the mechanical addition point 27, the control load F _c , which is the output of the charge-to-charge converter 34, the spring load F ₀ of the return springs 2A and 2B, and the brake. The braking torque equivalent load F _b added from the first housing 12 via the swing arm 13 is added. On the spool 20, the displacement x corresponding to the force balance of each load, which is the output of the mechanical addition point 27, is
20 via the inertial delay element 28 which corresponds to the resistance during movement, the displacement x including the hydraulic hose from the hydraulic control valve 2 to the brake 1 and the pressure with a first-order lag transfer function with gain K _p. It is converted into a brake oil pressure P _b in the converter 29 and is sent to the brake 1.

The gain K _p of the pressure converting unit 29 is a pressure gradient coefficient showing the rate of change of the brake oil pressure P _b with respect to the displacement x of the spool 20, and the rate of change with time of the displacement x of the spool 20 generated as described above. The (moving speed) is detected by the speed detector 42 and is fed back to the servo amplifier 5 as a speed feedback signal E ₀ .

The brake 1 which generates a braking torque T _b in response to the supply of the brake oil pressure P _b regulated by the hydraulic control valve 2 has a hydraulic pressure addition point 35 and a torque conversion portion as shown in the figure.
Replaced with 36. The hydraulic pressure addition point 35 is the brake 1
This is equivalent to the working cylinder 10 of
In the above, the brake hydraulic pressure P _b supplied from the hydraulic control valve 2 and the biasing force of the return springs 101, 101 ...
The equivalent piston return pressure P _r obtained by dividing by the pressure receiving area at 10a is added to the piston 100 of the working cylinder 10.
Presses the brake plate by the force multiplied by the difference between the two pressures (P _b -P _r) to the pressure receiving area.

This pressing force passes through the torque converting section 36 represented as a proportional element having a proportional sensitivity K _c , and the braking torque T.
converted to _b . The proportional sensitivity K _c of the torque converter 36
Is the braking torque T _{b with} respect to the pressing force of the piston 100.
Is a brake constant indicating the rate of change of the brake plate. This brake constant K _C is the effective diameter D _{m of the} brake plates and the number Z, the dynamic friction coefficient μ between the brake plates, the pressure receiving area A of the piston 100, and other dimensions of each part of the brake 1. And a constant that depends on the viscosity of the oil enclosed in the housing 12. T _b = K _c (P _b −P _r ) ∵K _c = μZD _m A

The braking torque T _b thus generated is the braking torque equivalent load F _b divided by the length L of the swing arm 13.
Is added to the spool 20 of the hydraulic control valve 2, that is, the mechanical addition point 27. Also, this braking torque equivalent load F
_b is a load cell 41 mounted on the supporting portion of the swing arm 13.
The detection result is sent to the servo amplifier 5 as a torque feedback signal E _b corresponding to the braking torque equivalent load F _b via the feedback amplifier 37. The feedback amplifier 37 is a fixed amplifier that amplifies the output of the load cell 41 in order to obtain the torque feedback signal E _b having a level corresponding to the torque command signal E _s .

FIG. 5 is a block diagram showing the internal structure of the servo amplifier 5. As shown in the figure, the servo amplifier 5 includes a PD computing unit 7 in which a proportional element 70 having a proportional sensitivity K _e and a differentiating element 71 having a gain K _d are arranged in parallel between a pair of adders 72, 73, and the PD computing unit 7. It comprises a feedforward circuit arranged in parallel in the arithmetic unit 7 and a series lead compensating element 54 arranged on the output side of the PD arithmetic unit 7 and having a first-order lead transfer function having a gain K ₀ .

The PD calculation unit 7 is a main part for determining the control amount of the pressure control valve 32, and the torque command signal E _s input to the servo amplifier 5 is given to the adder 72 of the PD calculation unit 7. A torque feedback signal E _b corresponding to the braking torque T _b generated by the brake 1 is also applied to the adder 72 via a filter 74, and the adder 72 outputs a deviation signal (= E _s −) between the two. E _b ) is output to the proportional element 70. The torque feedback signal E _b is directly applied to the differential element 71 arranged in parallel with the proportional element 70, and the output of the proportional element 70 is differentiated by the adder 73 on the output side of both elements 70 and 71. PD calculation for weight reduction correction is performed by the output of 71.

The feedforward circuit provided in parallel in the PD computing section 7 comprises a feedforward amplifier 50 which is a differential element having a gain K _f, and a non-linear processing section 51 on the output side thereof. Feed forward amplifier 50
Is the torque command signal E branched at the front side of the adder 72.
_s is input, and the output of the feedforward amplifier 50 corresponding to the rate of change of the torque command signal E _s is given to the adder 73 of the PD calculation unit 7 via the non-linear processing unit 51. As shown in the figure, the non-linear processing unit 51 does not output an output for a small input whose absolute value is smaller than a predetermined lower limit value,
The control amount of the pressure control valve 32 determined by calculation in the PD calculation unit 7 is configured to generate an output that proportionally increases / decreases with respect to an input exceeding the lower limit value, and the nonlinear amount applied to the adder 73 The output of the processing unit 51 is used to make an increase correction.

On the other hand, the detection result of the speed detector 42 input to the servo amplifier 5, that is, the speed feedback signal E ₀ corresponding to the moving speed of the spool 20 of the hydraulic control valve 2 is output by the amplifier 52 having the gain K _V. After the gain is adjusted, P
It is given to the adder 53 on the output side of the D calculation unit 7. The adder 53 is a PD calculation unit 7 corresponding to the control amount of the pressure control valve 32.
Obtains the output signal Z ₀ and the deviation between the speed feedback signal E _0, gives the deviation signal to the series lead compensator 54, the series lead compensator 54, the primary output signal Z ₀ of the PD computing section 7 Is output, and this output is given to the charge-to-charge conversion section 34 corresponding to the pressure control valve 32.

As described above, in the servo amplifier 5, the control amount of the pressure control valve 32 is based on the deviation between the torque feedback signal E _b and the torque command signal E _s corresponding to the braking torque T _b during the generation of the brake 1. It is determined by the PD computing unit 7 including a proportional element 70 that performs a proportional operation and a differential element 71 that performs a differential operation based on the torque feedback signal E _b , and this control amount is increased by the output of the feedforward circuit. To be done.

When the torque command signal E _s is changed, the output of the feedforward circuit immediately increases according to this change, while the brake 1 follows the change.
Since the output of the differentiating element 71, which is the differential value of the torque feedback signal E _b , is small before the braking torque T _b of the torque command signal E _s starts to change, the output of the PD computing unit 7 at the initial stage of changing the torque command signal E _s is the torque command. The output of the proportional element 70 corresponding to the deviation between the signal E _s and the torque feedback signal E _b is corrected to the increase side. Therefore, the control oil pressure P _c sent from the pressure control valve 32 to the oil pressure control valve 2 suddenly changes, and the brake oil pressure P _b sent to the brake 1 suddenly changes due to the movement of the spool 20 according to this change. Therefore, the time required to realize the braking torque T _b corresponding to the torque command signal E _s is shortened, that is, high responsiveness is achieved.

On the other hand, the braking torque T _b of the brake 1 changes in accordance with a sudden change in the brake oil pressure P _b , and the differential element
The output of 71 increases and the output of the PD calculator 7 is reduced and corrected. As a result, the brake oil pressure P _b sent to the brake 1 is increased.
The rate of change of (1) becomes gentle before the generated braking torque T _b approaches the target torque, the optimization of damping can be achieved, and the generation of harmful zero-cross peaks can be greatly reduced.

Furthermore, the control signal output from the servo amplifier 5 is passed through the series advance compensation element 54 rather than the output itself of the PD computing unit 7, so that the spool 20 of the hydraulic control valve 2 is controlled.
Since the signal is a signal obtained by adding a primary advance related to the velocity feedback signal E ₀ corresponding to the moving velocity of the above-mentioned output to the output, as described above, the pressure conversion unit 29 including the pressure control valve 32 and the hydraulic piping system has. The first-order lag component is absorbed. In particular, when the time constant T ₀ of the series lead compensation element 54 is set to be substantially equal to the time constant T of the pressure conversion unit 29, the primary delay of the hydraulic piping system is almost completely canceled out, and the responsiveness due to this delay is reduced. It can eliminate the deterioration.

In the load testing apparatus according to the present invention having the above-mentioned basic control system, the load test of the test rotating machine A linked to the brake 1 is performed at a constant speed or a constant power. Therefore, a torque command changing section 6 for continuously changing the torque command signal E _s is provided in parallel with the servo amplifier 5.
FIG. 6 is a block diagram showing a configuration of a control system including the torque command changing unit 6. In order to prevent the drawing from becoming complicated, in this figure, the entire basic control system shown in FIG. 4 is replaced with a transfer element 8 having a transfer function G (= T _b / E _s ). The transfer element 8 receives the torque command signal E _s given to the servo amplifier 5 and outputs the braking torque T _b generated by the brake 1, and the transfer function G is as shown in the following equation. A second-order lag system including the respective values in FIGS. 4 and 5 is obtained.

[0052]

[Equation 1]

The braking torque T which is the output of the transmission element 8
_b is a test rotary machine A connected to the rotary shaft 11 of the brake 1.
Be transmitted to. The test rotary machine A can be replaced with the mechanical addition point 80 and the inertia element 81, and the braking torque T _b generated by the brake 1 is added to the test rotary machine A at the mechanical addition point 80, so that the test rotary machine A The rotation speed is determined by the difference between the driving torque T _{0 of} itself and the braking torque T _b of the brake 1 via an inertial element 81 including the load inertia J ₀ of the rotating portion, and this rotation speed is the rotation axis 11 of the brake 1. Is detected by the rotation speed detector 43 connected to.

The rotation speed detector 43 is 30 / π, as shown in the figure.
By multiplication of the unit conversion constant 44 with proportional sensitivity
The rotation speed of the test rotating machine A having a speed dimension of N (rpm) is converted into a rotation signal having the number of pulses per unit time corresponding to this, and the rotation signal is detected. The output signal of the converter 43 is converted into a voltage signal by the F / V converter 82 having a conversion constant K _n, and the polarity at the time of reverse rotation is converted by the polarity converter 84 to obtain a level corresponding to the torque command signal E _s. The rotation speed signal E _n has a sign and is supplied to the torque command changing unit 6 which is a feature of the present invention.

The torque command changing unit 6 has the first and second PIs.
The arithmetic units 60 and 61 and the respective adders 62 and 63 on the input side thereof
And an analog memory 64 and an adder 65 which are interposed on the input side of the servo amplifier 5. The first PI calculator 60 used in the constant speed control includes a proportional element 60a having a proportional sensitivity A ₁ and an integral element 60b having a gain A ₂ arranged in parallel. the adder 62 acts torque command signal E _s as the target rotational speed command signal, the deviation between the rotational speed signal E _n is calculated, are no to give the PI computing unit 60.

On the other hand, the second used in constant power control
PI computing unit 61, the proportional element 61a having a proportional sensitivity A ₃
And an integrating element 61b having a gain A ₄ are arranged in parallel. The input side of the adder 63 of the of the PI computing unit 61, the torque command signal E _s, the output polarity converter multiplier 66 for multiplying said rotational speed signal E _n and the torque feedback signal E _b 67, and the adder 63 is
Torque command signal E acting as target absorption power command signal
_s and the output of the multiplier 66 resulting from the multiplication,
That is, the deviation from the absorbed power due to the braking of the brake 1 is obtained and output to the PI calculator 61.

The output sides of the PI calculators 60 and 61 are connected to the adder 65 via the selection switch SW ₁ . The selection switch SW ₁ has three selection positions of T, N and L as shown in the figure. When the selection switch SW ₁ is selectively operated to the N position, the output of the first PI calculator 60 is ,
Similarly, when the L position is selected and operated, the output of the second PI calculator 61 is supplied to the adder 65, respectively. On the other hand, when the selection switch SW ₁ is in the T position, the torque command signal E _s is directly given to the adder 65.

The analog memory 64 holds the torque command signal E _s at that time in response to turning on of the associated memory switch SW ₂ , and continuously outputs the held value to the adder 65. . Memory switch SW ₂ is
The selection switch SW ₁ is interlocked with the selection switch SW _{1.
When 1} is in the N position or L position, it is turned on, and when it is in the T position, it is turned off.

In the constant torque control described above, the selection switch S
It is implemented with W ₁ as the T position. At this time, since the memory switch SW ₂ is turned off and the analog memory 64 is in a non-operating state, the adder 65 does not perform any addition,
The torque command signal E _s is directly applied to the servo amplifier 5, and the brake 1 is operated by the operation of the servo amplifier 5 described above.
Is the braking torque T _b corresponding to the torque command signal E _s.
Is continuously generated, and the load characteristic under the braking torque T _b is obtained.

On the other hand, when the selection system SW ₁ is selectively operated to the N position, the analog switch 65 holds the torque command signal E _s at the present time by the ON operation of the memory switch SW ₂ which is interlocked therewith, and the servo amplifier 5 Block direct input to. Further, by the selection operation of the N position, the output of the first PI calculator 60 is given to the adder 65, and the input to the servo amplifier 5 is the torque command signal E _s , that is, the target rotation speed command. a PI calculation value of the deviation between the rotational speed signal E _n of the signal and the test試回turning point a, the holding value of the analog memory 65, i.e., becomes a deviation signal between the torque command signal E _s at the switching point, the servo amplifier 5 As a result of performing the control operation with respect to this deviation signal, it becomes possible to perform a load test while maintaining the rotation speed before the switching, and a load characteristic at a constant speed can be obtained.

[0061] or more as the constant speed control is performed, and inputs the torque command signal E _s supplied to the torque command change section 6, made as an output rotational speed E _n of the test試回turning point A,
Transfer function G ₁ (= E _n / E _s ) of the entire control system in this case
By deriving, the following quadratic delay system similar to the transfer function G during constant torque control can be arranged. It should be noted that, here, the time variation between the driving torque T ₀ of the test rotating machine A and the held value of the analog memory 64 is zero, and the transfer function G shown in the equation (1) has a low ω _0. It is assumed that the proportional constant K can be substituted under the condition of being in the frequency range.

[0062]

[Equation 2]

In the constant speed control as described above, the test rotating machine A is started, the desired torque command signal E _s is given to obtain the constant torque state, and then the selection switch SW _{1 is set} to the N position to set the PI calculator. Proportional Sensitivity A ₁ and Integral Element of Proportional Element 60a in 60
This is carried out by the procedure of optimally adjusting the gain A _{2 of} 60b.

Similarly, the load test under constant power is performed by operating the selection switch SW ₁ to the L position during the execution of constant torque control under the appropriate torque command signal E _s . In this case, similarly to the case of the constant speed control, the analog memory 65 holds the torque command signal E _s at the present time by the ON operation of the memory switch SW ₂ which is interlocked with the selection switch SW ₁ , and the servo amplifier 5 directly The input of the second PI calculator 61 to the adder 65,
The input to the servo amplifier 5 is the torque command signal E _s , that is, the PI calculation value of the deviation between the target absorption power command signal and the absorption power due to the braking of the brake 1, and the value held in the analog memory 65, that is, the switching time point. Is a deviation signal from the torque command signal E _s, and as a result of the servo amplifier 5 performing a control operation for this deviation signal, it becomes possible to carry out a load test while maintaining the absorbed power before the switching, and under constant power. Load characteristics can be obtained.

The constant power control performed as described above is performed with the torque command signal E _s given to the torque command changing unit 6 as an input, and the absorbed power by the brake 1 corresponding to the output of the multiplier 66 as the final output. When the transfer function G ₂ (= L / E _s ) of the entire control system in the case is derived, it can be rearranged into a quadratic delay system as shown in the following equation, similarly to that in the constant speed control. Note that here, the transfer function G expressed by the equation (1) is
Makes the assumption that ω ₀ can be replaced with the proportional constant K under the condition that ω ₀ is in the low frequency range, and in the state of constant power operation, T ₀ −T _b > 0, and further T ₀
It is assumed that the assumption that ≈T _b holds.

[0066]

[Equation 3]

The output of the first PI calculator 60 used in constant speed control is given to the adder 65 as a positive feedback signal, whereas the output of the second PI calculator 61 used in constant power control is constant. The output is a negative feedback signal because it is necessary to increase the braking torque T _b of the brake 1 as the rotation speed of the test rotating machine A increases in order to maintain the constant speed, This is because the braking torque T _b needs to be reduced as the absorbed power of the brake 1 increases in order to maintain the above.

[0068]

As described above in detail, in the load testing apparatus according to the present invention, the input to the servo amplifier in the constant speed control is performed by the PI calculation based on the deviation between the rotation speed of the rotating machine under test and the torque command signal. Also, the input to the servo amplifier in the constant power control is determined by the PI calculation based on the deviation between the absorbed power by the brake and the torque command signal, and the brake hydraulic pressure is adjusted based on these input signals. By the configuration that determines the controlled variable for
The transfer function of the entire control system in constant speed control and constant power control becomes a quadratic delay system like that in constant torque control, and the response of the braking torque generated by the brake according to the control signal from the servo amplifier is It is high, and it becomes possible to carry out a constant speed test and a constant power test with high accuracy in order to respond to the load condition when using the test rotating machine.
The present invention has excellent effects.

[Brief description of drawings]

FIG. 1 is a schematic diagram showing a configuration of a hydraulic circuit of a load test apparatus according to the present invention.

FIG. 2 is a vertical cross-sectional view of a hydraulically actuated brake that is a means for generating load torque.

FIG. 3 is an enlarged view of a main part of FIG.

FIG. 4 is a block diagram showing an overall configuration of a control system in constant torque control of the load test apparatus according to the present invention.

FIG. 5 is a block diagram showing an internal configuration of a servo amplifier.

FIG. 6 is a block diagram showing a configuration of a control system including a torque command changing unit, which is a feature of the present invention.

[Explanation of symbols]

1 Brake 2 Hydraulic control valve 5 Servo amplifier 6 Torque command change part 10 Working cylinder 10a Oil chamber 10b Oil chamber 11 Rotating shaft 12 Housing 13 Swing arm 32 Pressure control valve 41 Load cell 42 Speed detector 43 Rotation speed detector 60 PI calculation vessel 61 PI calculator 64 analog memory 82 F / V converter 100 piston A test試回turning point E _s torque command signal E _b torque feedback signal E _n rotational speed signal

Claims (2)

[Claims] 1. A hydraulically actuated brake, which is interlocked with a rotating machine under test, a pressure adjusting means of a brake hydraulic pressure supplied to the brake, a feedback signal of a braking torque generated by the brake, and an externally applied feedback signal. A servo amplifier that controls the pressure adjusting unit based on a torque command signal, and a torque command changing unit that adjusts the torque command signal according to a change in the rotation speed of the test rotating machine are provided. In the load test apparatus capable of performing the load test of the rotating machine under test at a constant speed by giving the output of the unit to the servo amplifier, the torque command changing unit includes the torque command signal and the rotation speed. Of the detection signal and the input of a PI calculator, an analog memory that holds a reference average target torque, and a deviation of the outputs of both of them, A load test device, comprising: an adder provided to a voamp. 2. A hydraulically actuated brake, which is interlocked with the rotating machine under test, a means for adjusting the brake hydraulic pressure supplied to the brake, a torque command signal given from the outside, and a braking torque generated by the brake. Obtained as a product of a servo amplifier for controlling the pressure adjusting means based on the feedback signal and a feedback signal of the braking torque and a detection signal of the rotation speed of the rotating machine under test, and the absorption power of the brake. A torque command changing unit that adjusts the torque command signal according to a change is provided, and by applying the output of the torque command changing unit to the servo amplifier, the load test of the test rotating machine can be performed under constant power. In the load test apparatus configured as described above, the torque command changing unit receives the deviation P between the torque command signal and the multiplication value as the input P.
A load test apparatus comprising: an I calculator, an analog memory that holds a reference average target torque, and an adder that obtains a deviation between the outputs of the two and outputs it to the servo amplifier.