JPH0726883B2 - Load test equipment - Google Patents

Description

TECHNICAL FIELD The present invention relates to a load test apparatus used for load testing of rotating machines such as reduction gears, and more specifically, generation of a braking torque applied to a test rotating machine. The present invention relates to a load test device using a hydraulically actuated brake as a means.

[Conventional technology]

A load test device using a hydraulically actuated brake as a braking torque generating means is disclosed, for example, in Japanese Utility Model Application No. 62-46342.
It is disclosed in Japanese Patent Publication No. The brake is a so-called wet multi-plate brake, and includes a housing supported in a fixed state and a rotary shaft rotatably supported in the housing.
Alternately, a large number of braking plates coaxially attached to the outside of the rotary shaft while being constrained from rotating and a large number of braking plates similarly attached to the inside of the housing via oil filled in the housing. The braking torque is generated by the shear resistance of the oil film between the respective braking plates. The load test device using this brake, the housing,
From this, a swinging arm having a predetermined length that projects radially outward is swingably supported within a predetermined angle range.
In addition, an operating cylinder that operates by hydraulic pressure and presses the braking plates on the housing side and the shaft side in the axial direction is provided inside the housing, while supplying hydraulic pressure to the operating cylinder at the supporting position of the swing arm. A hydraulic control valve for controlling is provided, and the test rotary machine is used in conjunction with the rotary shaft. The hydraulic control valve includes a valve element that is biased in one direction in advance with a force equivalent to a set torque during a load test, and the swing arm is connected to this valve element to generate a braking force. Torque acts on the valve body via the arm. The urging force of the valve body is also hydraulic pressure, and the urging force is changed by, for example, changing the distribution of control hydraulic pressure to be fed to two oil chambers formed on both sides of the valve body. At this time, a load cell and a speed detector interlocking with the valve body are provided at the connecting portion between the valve body and the swing arm, and the detection result of the braking torque equivalent load by the former and the displacement speed of the valve body by the latter are detected. Is used as a feedback signal.

Therefore, first, the working cylinder advances by the working oil fed through the hydraulic control valve in a state in which the valve body is biased by the hydraulic pressure corresponding to the set torque, and thereby the working cylinders mutually move. Braking torque is generated by the shearing of the oil film between the pressed braking plates. This braking torque acts on the valve body of the hydraulic control valve via the housing and the swing arm, and when this acting force exceeds the hydraulic pressure, the displacement of the valve body causes the working cylinder to retract and the braking torque Is canceled. That is, the valve body of the hydraulic control valve is displaced according to the balance between the braking torque equivalent load and the biasing force applied to the valve body in advance, and this displacement controls the continuous supply of the hydraulic oil to the working cylinder. As a result, the braking torque is balanced near the set torque, and the load test of the test rotating machine at the set torque is performed.

[Problems to be Solved by the Invention]

Now, in a load test performed using such a load test apparatus, in order to simulate various load states during actual use of the test rotating machine, a test in which the set torque is changed at any time is desired. However, in order to enable frequent changes in the load, first, the braking torque actually generated by the brake is required to respond to changes in the set torque at a high speed. The applicant of the present application makes the working cylinder a double-acting type, and allows the piston of the working cylinder to be forcibly displaced not only on the braking side but also on the releasing side.
It is possible to follow the change of the set torque at an extremely high speed,
A load test apparatus that can meet the above requirements was proposed in Japanese Utility Model Application No. 63-100060. By the way, the load torque simulated in the actual load test is the number of revolutions (including power).
Rotation speed, such as a load torque that changes in proportion to changes in the load, a constant load torque regardless of the change in the rotation speed, and a load torque such that the power, which is the product of the rotation speed and the load torque, becomes constant. However, the load test apparatus proposed in Japanese Patent Application No. 63-100060 was not able to selectively realize such a load state.

The present invention has been made in view of such circumstances, and the load test apparatus proposed in Japanese Utility Model Application No. 63-100060 is developed to provide a test rotating machine under a load condition that changes in relation to the rotation speed and the torque. It is an object of the present invention to provide a load test device capable of performing the load characteristic test of the above.

[Means for Solving the Problems]

A load test apparatus according to the present invention includes a hydraulically-operated brake in which a rotary shaft that is interlockingly connected to a test rotating machine is housed in a swingably supported housing, and the brake is braked and released. A double-acting actuating cylinder for moving the piston in each direction is provided, while a preset torque equivalent force that is configured in the bearing portion of the housing and that is applied in the opposite direction to the housing and the swing arm. A load test device comprising a hydraulic control valve for controlling the hydraulic pressure to be fed to the working cylinder by the displacement of the valve body caused by the difference between the force applied via the load test device and the rotational speed of the rotary shaft. A rotation speed detector and means for continuously changing the set torque equivalent force applied to the valve body of the hydraulic control valve based on the detection result, means for detecting a force applied via the swing arm, and the means And means for electrically correcting the detected result set torque equivalent force based on the means for detecting,
It is characterized by having constant speed control, constant power control, and speed proportional torque control functions on the basis of constant torque control.

[Action]

In the present invention, the rotation speed of the rotating machine under test is detected as the rotation speed of the rotating shaft of the brake interlocked with this, and the hydraulic control valve for controlling the hydraulic pressure fed to the working cylinder based on the detection result is detected. The biasing force applied in advance to the valve element is changed, and the braking torque generated by the brake changes the load torque that changes in proportion to the rotation speed (including power), and the constant load torque regardless of the change in the rotation speed,
A load state that changes in relation to the rotation speed and the torque is realized, such as a load state in which the absorbed power that is the product of the rotation speed and the braking torque is constant.

〔Example〕

Hereinafter, the present invention will be described in detail with reference to the drawings showing an embodiment thereof.

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 longitudinal sectional view showing an internal structure of a hydraulically actuated brake used as a control torque applying means, and FIG. It is a principal part enlarged view of FIG.

The load test apparatus according to the present invention includes a hydraulically operated brake 2 that generates a braking torque to be applied to a test rotating machine, and an operating cylinder 10 that is configured therein to perform braking and releasing operations of the brake 2. A hydraulic control valve 6 for controlling the hydraulic pressure supplied to the operating cylinder 10, a hydraulic pump 9 for generating the operating hydraulic pressure of the operating cylinder 10, a hydraulic pump 50 for generating the control hydraulic pressure of the hydraulic control valve 6, and the hydraulic pump 50. And a hydraulic pressure control valve 6, which are provided between the pressure control valve 6 and a current control type pressure control valve 12 for distributing the control hydraulic pressure.

First, the structure of the brake 2 will be briefly described with reference to FIG. The brake 2 is provided with a rotary shaft 1 that is interlocked with a rotary shaft A of the test rotating machine, and a housing 3 that rotatably supports the rotary shaft 1 and oil is enclosed in the rotary shaft 1. Is a wet multi-disc brake that generates a braking torque due to the shear resistance of the oil film with the rotation of the rotating shaft 1 between a large number of braking plates mounted with restraint. In FIG. 2, the upper half is in the braking state and the lower half is in the released state.

The rotary shaft 1 is properly inserted into the housing 3 from the left side of the figure, and the rotary shaft A side, that is, the left side of the figure, has bearings 27A and 27B.
, And the tip side, that is, the right side in the figure, is rotatably supported by bearings 27C, and the housing 3 together with the rotating shaft 1 is a pair of bearing bases 3 fixed on both left and right sides.
Between 0 and 30, it is rotatably supported via separate bearings 28A and 28B. Inside the housing 3, the rotary shaft 1
A surrounding cylinder 91 is provided to surround the outer side of the inside of the surrounding cylinder 91. Inside the surrounding cylinder 91, rotation is restrained and a large number of braking plates are attached so as to be movable in the axial direction. Further, also in the portion surrounded by the surrounding cylinder 91 on the outside of the rotary shaft 1,
Rotation is restricted by the rotary shaft 1, and a large number of braking plates are attached so as to be movable in the axial direction, and both braking plates are alternately superposed in the axial direction. A rotary joint 92 is rotatably fitted to the right end portion of the housing 3 via a bearing 28C, and the rotary joint 92 is housed in the surrounding cylinder 91.
Oil introduced and discharged through 92 and an oil passage 94 connected to this is enclosed. The left end of this oil passage 94 is the rotary shaft 1
Most of the oil introduced through the oil passage 94, which is connected to the oil passage 1a formed at the center of the bearing 27A and 27B, is used for lubricating and cooling the brake plates. But bearings 27A, 28B and high pressure oil seal 27D
It is also used as a lubricating oil.

The brake 2 generates braking torque by moving the control plates close to each other in the surrounding cylinder 91 and releases the braking torque by moving them away from each other, and the braking torque is released on the right side of the surrounding cylinder 91. An operating cylinder 10 for pressing a brake plate to perform the approach is constructed. The working cylinder 10 is of the double-acting type proposed in Japanese Utility Model Application No. 63-100060, and has two oil chambers 10a and 10b,
When the hydraulic oil is supplied to one of the oil chambers 10a, the piston 100 moves forward to press the braking plate closer, and when the hydraulic oil is supplied to the other oil chamber 10b, the piston 100 is On the contrary, it is designed to move in and out. Further, as shown in FIG. 3, the pistons 100 are quickly separated from each other due to the retreat of the piston 100, between the braking plates.
Return springs 103, 103 ... Which urge these members in a separating direction to each other are provided so as to create a gap.

The working cylinder 10 has a stepped ring-shaped piston 100 fitted in a stepped ring hole (a large diameter on the braking plate side and a small diameter on the side farther from the stepped plate) formed around the right end portion of the surrounding cylinder 91. In the oil chamber 10a formed on the right side of the large diameter portion of the stepped annular hole, the oil passage 93 formed on the right side surface of the housing 3 is provided.
The hydraulic oil is introduced via a and into the oil chamber 10b formed on the left side of the small diameter portion via an oil passage 93b similarly formed.

An oscillating arm 4 of a predetermined length is provided on the outside of the housing 3 of the brake 2, and the oscillating arm 4 is supported via a valve body 5 of a hydraulic control valve 6 fixed in a vertical position. Cage,
A load cell 18 for detecting a load equivalent to a braking torque applied to the valve body 5 via the arm 4 is provided at a connecting portion between the valve body 5 and the swing arm 4. The valve body 5 has four large-diameter portions arranged in parallel in the axial direction at predetermined intervals,
The hydraulic control valve 6 includes a first oil chamber 61 formed on the lower side of the large diameter portion and a second oil chamber formed between the large diameter portions in order from the lower side.
An oil chamber 62, a third oil chamber 63, a fourth oil chamber 64, and a fifth oil chamber 65 formed on the uppermost side of the large diameter portion are provided. The oil supplied to the working cylinder 10 generated by the hydraulic pump 9 is supplied via an accumulator to a pump port 63a which is always in communication with the third oil chamber 63 regardless of the displacement of the valve body 5, and The control pressure oil generated by the hydraulic pump 50 is connected to the control port 61a which is always connected to the first oil chamber 61 via the separate control ports 39A and 39B of the pressure control valve 12 and the fifth control port 61a. The oil is supplied to the control port 65a that is always connected to the oil chamber 65. In addition, the second oil chamber 62 and the fourth oil chamber 64 are provided with separate return ports 62 that are always connected to them.
It is opened to the oil tank T maintained at a low pressure via a and 64a. Return springs 7A and 7B for urging the valve body 5 toward the center are arranged in the first oil chamber 61 located at the bottom and the fifth oil chamber 65 located at the top, respectively. When there is no pressure difference between the oil chambers 61 and 65, the valve body 5 is stopped at a position (neutral position, spring center) determined by the urging force of the return springs 7A and 7B.

Further, in the hydraulic control valve 6, when the valve body 5 is in the neutral position, the hydraulic control valve 6 is closed at the large-diameter portions on both sides of the third oil chamber 63.
A pair of brake ports 11R and 11F are opened. now,
When in the forward rotation direction (the direction in which the compressive load is applied to the load cell 18), the hydraulic equilibrium state between the first oil chamber 61 and the fifth oil chamber 65 is broken, and the valve element 5 is moved upward (or downward). Brake port 11 located on the upper side (or lower side) when displaced to
F (or 11R) opens into the third oil chamber 63, and the working oil supplied into the third oil chamber 63 passes through a 4-port 3-position switching type electromagnetic switching valve 16 and one of the working cylinder 10 Oil chamber 10a
(Or 10b), and at this time, the other brake port 11R (or 11F) is connected to the second oil chamber 62 (or the fourth oil chamber).
64) and the return port 62a (or the return port 64a)
The other oil chamber 10b (or 10a) of the working cylinder 10 is opened to the oil tank T via the, and is directly communicated with the oil tank T. The electromagnetic switching valve 16 is provided to prevent the brake lock by switching the flow path of the hydraulic oil according to the forward and reverse of the rotation direction of the rotary shaft 1.

In order to detect the velocity at the time of displacement of the valve body 5 thus generated, the portion of the valve body 5 integrated with the cylinder block is
A speed detector 19 which is a linear speed type converter is connected, and the speed detector 19 generates a speed signal of a level proportional to the displacement speed of the valve body 5. A rotation speed detector 17 for detecting the rotation speed of the rotary shaft 1 is fixedly mounted on the support base 30 located on the protruding side of the rotary shaft 1 in the brake 2.
The rotation of the rotation shaft 1 is transmitted to the rotation speed detector 17 by a timing belt 17a wound between the rotation speed detector 17 and the protrusion of the rotation shaft 1, and a predetermined number of pulses per rotation of the rotation shaft 1 are transmitted. A rotary encoder or the like that generates a signal is used. The detection results by the speed detector 19 and the rotation speed detector 17, together with the detection result of the load equivalent to the control torque of the brake 2 by the load cell 18, serve as a feedback signal when changing the set torque, specifically, the pressure control valve.
It is used to control 12 operating currents.

The load testing apparatus having the above-described configuration applies a signal corresponding to a set torque, which will be described later, to the pressure control valve 12 and uses the electromagnetic switching valve 16 in a state in which the electromagnetic switching valve 16 is switched according to the rotation direction of the test rotating machine. The control hydraulic pressure generated by the hydraulic pump 50 is distributed by the pressure control valve 12, and then the control port 61a connected to the first oil chamber 61 and the control port 65a connected to the fifth oil chamber 65.
The pressure difference corresponding to the set torque is generated between the oil chambers 61 and 65. This pressure difference is, for example,
When the first oil chamber 61 is generated toward the fifth oil chamber 65, the valve body 5
Along with the upward movement of the brake port 11
F is in the third oil chamber 63, and brake port 11R is in the second oil chamber.
Since each of the openings 62 opens, the working oil supplied into the third oil chamber 63 via the pump port 63a is supplied to the advancing side oil chamber 10a of the working cylinder 10 via the electromagnetic switching valve 16, and At the same time, the hydraulic oil staying in the entry-side oil chamber 10b is quickly returned to the oil tank T via the electromagnetic switching valve 16 and the second oil chamber 62 of the hydraulic control valve 6. As a result, the piston 10
0 rapidly advances, the braking plates are pressed against each other, the braking resistance is generated by the brake film due to the shear resistance of the oil film between the braking plates, and the housing 3 is not rotated in the rotation direction of the rotating shaft 1. And This braking torque becomes a braking torque equivalent load through the swing arm 4 and is applied to the valve body 5, and the valve body 5 is pressed downward by this applied force, so that the first oil chamber 61 and It stops at a position where the pressure difference between the fifth oil chambers 65 and the braking torque-equivalent load applied via the swing arm 4 are in equilibrium.

Now, when the set torque changes, the valve body 5 moves according to the change in the control oil pressure, which causes the brake port 11R or 11F to move into the third oil chamber 63 or the brake port 11F or As a result of the 11Rs opening into the fourth oil chamber 64 or the second oil chamber 62, respectively, the hydraulic oil is quickly supplied to one of the oil chambers 10a and 10b of the working cylinder 10 and discharged from the other. Done. Accordingly, the piston 10 of the working cylinder 10
0 indicates that the valve body 5 moves at high speed to advance or retreat, and the brake 2 is quickly braked and released.
The valve body 5 reaches a new equilibrium state and stops, and the brake 2
Will generate a braking torque corresponding to the new set torque thereafter. That is, in this load testing device, the double-acting actuation cylinder 10 and the hydraulic control valve 6 having the above-described configuration allow the brake 2 to be quickly braked and released with respect to changes in the set torque, which is extremely favorable. Responsiveness is realized.

The operation of the load test apparatus according to the present invention has been described above by paying attention to the oil flow. In this apparatus, the load equivalent to the braking torque generated by the brake 2 is moved by the load cell 18 and the load is added. The displacement speed of the valve body 5 of the hydraulic control valve 6 is detected by the speed detector 19, and the rotation speed of the rotating shaft 1 of the brake 2 that rotates while being further applied with the braking torque is detected by the rotation speed detector 17. Therefore, feedback control is performed based on these detection results. In particular, the feature of the present invention is that the set torque equivalent voltage in the brake 2 can be continuously changed based on the detection and calculation results of the detectors.

FIGS. 4-a and 4-b are block diagrams of the control system of the load test apparatus according to the present invention, and include all of the electrical system, the hydraulic system, the mechanical system, and the rotating machine under test.

First, the constant torque control function, which is a basic control function, will be described. E in FIG. 4-a is the target torque equivalent voltage set by the torque setting device (see FIG. 6). This is equal to the set torque equivalent voltage E _s (E = E _s ) when the switches SW ₁ and SW ₂ are both off, and the servo amplifier
Entered in 80. In the servo amplifier 80, this set torque equivalent voltage E _s is first given to the addition point 81 and becomes a deviation signal (E _s −E _b ) from the feedback signal E _b corresponding to the detection result by the load cell 18, and the amplifier 28 Predetermined constant K
After being multiplied by ₁ , it is given to the adder 83 and becomes a deviation signal (= E _s −E _b −E _o ) from the feedback signal E _o corresponding to the detection result by the speed detector 19, and further is fed to the amplifier 84. Then, it is multiplied by a predetermined constant K ₂ and converted from current to hydraulic pressure and output.

Due to the electro-oil conversion of the control oil pressure of the pressure control valve 12 accompanied therewith, the output acts on both end faces of the valve body 5 and is converted into the dimension of force. It is converted into a dimension of force by 801 in FIG. 4 and input to the mechanical addition point 802. This addition point 8
02 corresponds to the valve body 5 of the hydraulic control valve 6, and the input to the addition point 802 is the hydraulic pressure applied to the valve body 5 by the control hydraulic pressure fed from the pressure control valve 12 and the valve end area. Is the product of With this force, the return spring 7A or 7B is applied to the valve body 5.
The urging force F _o by the brake force and the braking torque equivalent load F _b transmitted from the housing 3 of the brake 2 via the swing arm 4 are added, and the output of the addition point 802, that is, each of the above forces is applied to the valve body 5. The displacement x is generated through the multiplication of the algebraic sum of x and a predetermined delay element 803. In 804, this displacement x is multiplied by the pressure gradient constant K _p , converted into pressure, and input to the addition point 805.

The speed detector 19 detects the time rate of change (speed) of the displacement x, and K _{6 in} the figure is the speed detector.
19 shows the conversion constant of 19, and the output signal of the speed detector 19 obtained by such conversion is gain and phase adjusted through the phase compensation circuit 85 inside the servo amplifier 80, and is added as a speed feedback signal E _o . Is given to 83. Incidentally, as shown in FIG. 4-b, the same control can be achieved by a configuration in which the feedback circuit by the speed detector 19 in FIG. 4-a is eliminated and the amplifier 82 is replaced with the phase compensation circuit 82 '. It

The addition point 805 corresponds to the piston 100 of the working cylinder 10, and the input to the addition point 805 is the pressure of the working oil sent to the working cylinder 10 via the hydraulic control valve 6 and the electromagnetic switching valve 16. Is. In addition to this operating oil pressure, the piston 100 is applied with an equivalent piston return pressure P _r obtained by dividing the biasing force of the return spring 103 by the pressure received in the oil chamber 10a. The braking plate is pressed by the force generated by receiving the pressure corresponding to the pressure difference. As a result, the brake 2 generates the braking torque T _b obtained by multiplying the output of the addition point 805, that is, the difference between the two pressures, by a predetermined brake constant K _c at 806. Valve body 5 through swing arm 4
Applied force, i.e. the braking torque equivalent load F _b applied to the summing point 802 is obtained by dividing the braking torque T _b at the length l of the oscillating arm 4.

The load cell 18 detects the braking torque equivalent load F _b, and the signal output by the load cell 18 based on this detection is sent to the feedback amplifier 18a to set the torque equivalent voltage E _b.
A feedback signal E _b is obtained which is amplified to a level corresponding to _s and is given to the adder 81 of the servo amplifier 80. The above is the details of the constant torque control as the basic control function.

Next, the speed proportional torque control function will be described. The braking torque T _b generated by the brake 2 and the driving torque T _o applied from the test rotating machine side are added at an addition point 71, and further converted into a rotation speed of the rotary shaft 1 via an inertia element 72. This rotation speed is detected by the rotation speed detector 17, and the pulse signal output from the rotation speed detector 17 in response to this detection corresponds to the magnitude of the rotation speed by the F / V converter 73. Converted into a level analog rotation speed signal E _n . The rotation speed of the test rotating machine, that is, the rotating shaft 1 is applied to the brake 2.
The order to generate a braking torque T _b which changes corresponding to the rotational speed, utilizing the rotational speed signal E _n.

The rotational speed signal E _n is applied to the load condition setting unit 40, 41. The load condition setting unit 40 includes a power calculator 40a, a control mode changeover switch 40b, and a polarity converter 40c (interlocked with the rotation direction discriminator 41c). The changeover switch 40b is
It has three switching positions n ₁ , n ₂ and n ₃
E _n is given directly to n ₁ and is given to n ₂ and n ₃ after passing through the power calculator 40a. The output side of this the control mode switch 40b, it is the rotational speed signal E _n if the switching position n ₁ is selected, and the rotational speed signal E _n if the switching positions n ₃ is selected When a signal corresponding to the cube of the power of n ₃ is further selected, a signal corresponding to the n (n = 1 to 3) power of the rotation speed signal E _n is output. These signals are given to the adder 70 only when the selection switch SW ₁ is turned on after the polarities are converted by the polarity converter 40c according to the rotation direction. Therefore, the selection switch SW ₁ is turned on and the load condition setting unit 4
When 0 is selected, a feedback signal corresponding to the n-th power (1 ≦ n ≦ 3) of the rotation speed of the rotary shaft 1 is given to the adder 70 according to the switching position of the changeover switch 40b.
In the case of speed proportional torque control, the target value equivalent voltage E = 0
As a result of the new set torque equivalent voltage E _s being E _s = E ₁ being determined, the brake 2 is to generate the braking torque T _b that changes in proportion to the nth power of the change in the rotation speed. Become. This allows a load test under a load condition that changes in proportion to the nth power of the rotation speed.

Further, the constant speed control function will be described. Load condition setting section
41 is a multiplier 41a, a changeover switch 41b, a rotation direction discriminator 41
It is provided with c and a polarity converter 41d which works together with c. The selector switch 41b has three switching positions L, T, has N, the switching position N, the rotational speed signal ± E _n is given through the rotation direction discriminator (polarity converter) 41c There is. Further, the multiplier 41a is provided with a rotation direction discriminator 41c which is input to the multiplier 41a.
Is multiplied by the output of the feedback amplifier 18a (feedback signal E _b ) to output the increase / decrease in the power absorbed by the braking of the brake 2. This output is output in the direction of rotation by the polarity converter 41d. The polarity is converted accordingly and then applied to the switching position L. Therefore, on the output side of the load setting unit 41, when the changeover position of the changeover switch 40b is N, the analog memory 69 first holds the value of the target value setting voltage E, and the target value setting voltage E and the analog memory 69 are The input is cut off, and the polarity of the signal corresponding to the rotation speed of the rotary shaft 1 is converted according to the rotation direction.
The output signals are output as appropriate.
The speed deviation is detected at 42, and after being subjected to gain, phase adjustment, meter support calibration, etc., it is given to the adder 70 only when the selection switch SW ₂ is turned on. The changeover switch 41
No signal is given to the switching position T of b, and when the switching position T is selected, the output of the load setting unit 41 is always at the 0 level. Therefore, no addition is made in the adder 70, and the set torque equivalent voltage E
_s is equal to the target value equivalent voltage E, and the brake 2 operates to generate the braking torque T _b corresponding to the target value equivalent voltage E. When the switching position is switched to the N side from this state, a signal corresponding to the deviation between the rotation speed of the rotary shaft 1 before the switching and the current rotation speed is gained and phase-adjusted in the adder 70, and negative feedback is performed. Given as a signal, adder 70
This signal gives the servo amplifier 80 a set torque voltage E _s obtained by subtracting this signal from the value held in the analog memory 69. The compensation circuit 42 increases (or decreases) the current rotation speed.
In such a case, a negative (or positive) output corresponding to the deviation is issued. Therefore, when the rotation speed of the rotary shaft 1 increases (or decreases), the level of the set torque signal E _s increases (or decreases), and the brake 2 is generated in response to this, and is applied to the rotary shaft 1 of the brake 2. Since the braking torque _Tb also increases (or decreases), the rotation speed of the rotating shaft 1 decreases (or increases) until the deviation is eliminated. That is, when the changeover position of the changeover switch 41b is set to N, the load characteristic test can be carried out while variously changing the load torque while maintaining the rotation speed before the changeover.

Finally, the constant power control function will be described. In the same manner, when the changeover position of the changeover switch 41b is set to L, it is possible to carry out a characteristic test of constant load power by variously changing the load torque while maintaining the power generated by the rotating machine under test before this switching.

FIG. 5 shows the braking torque obtained by the execution of each control function described above, and FIG. 6 shows an example of the operation panel for selecting each function described above.

〔The invention's effect〕

As described above in detail, the load test apparatus according to the present invention is provided with the rotation speed detector that detects the rotation of the hydraulically actuated brake that is the braking torque generating means in association with the rotation axis of the brake. Based on the above, since there is a means for continuously changing the set torque equivalent force applied to the valve body of the hydraulic control valve for controlling the hydraulic pressure fed to the double-acting working cylinder of the brake, the rotation speed ( Various load states such as load state that changes in proportion to (including power), constant torque load state regardless of changes in rotation speed, constant rotation speed load state, constant power load state, etc. Therefore, the load characteristic test can be performed by applying a load corresponding to the actual load state to the test rotating machine. Further, since there is a means for electrically detecting the force equivalent to the set torque on the basis of the detection results of the means for detecting the force applied via the swing arm and the means for detecting the force, fine adjustment is possible and the damping force The present invention has excellent effects such as easy optimization and improved responsiveness.

[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 sectional view of a hydraulically actuated brake which is a means for generating a braking torque, and FIG. 4A and 4B are block diagrams of a control system of the load testing apparatus according to the present invention, and FIG. 5 is a graph showing braking torque obtained by implementing each control function described above. Sixth
The figure is a plan view of an operation panel for selecting each control function. 1 ... Rotary axis, 2 ... Brake 3 ... Housing, 4 ... Swing arm 5 ... Valve element, 6 ... Hydraulic control valve 10 ... Working cylinder, 10a, 10b ... Oil chamber 17 ... Rotation Speed detector, 18 load cell 19 Speed detector, 40, 41 Load condition setting section 100 Piston

Claims (1)

[Claims] 1. A hydraulically actuated brake in which a rotary shaft interlockingly connected to a test rotary machine is housed in a swingably supported housing, and a piston for braking and releasing the brake. A double-acting actuating cylinder for moving each is provided, and a force equivalent to a preset torque that is configured in the bearing portion of the housing and a force applied in the opposite direction to the force via the housing and the swing arm. A load testing device comprising a hydraulic control valve for controlling the hydraulic pressure to be fed to the working cylinder by the displacement of the valve body caused by the difference between the rotational speed detector and the rotational speed of the rotary shaft. And means for continuously changing the set torque equivalent force applied to the valve body of the hydraulic control valve based on the detection result, means for detecting the force applied via the swing arm, and the detection result based on the means. And a means for electrically correcting the set torque equivalent force, and based on constant torque control, each of the control functions of constant speed control, constant power control and speed proportional torque control is provided. Test equipment.