CN113605884B - Method for determining torsion shaft parameters of swing valve pulse generator - Google Patents
Method for determining torsion shaft parameters of swing valve pulse generator Download PDFInfo
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- CN113605884B CN113605884B CN202110808708.XA CN202110808708A CN113605884B CN 113605884 B CN113605884 B CN 113605884B CN 202110808708 A CN202110808708 A CN 202110808708A CN 113605884 B CN113605884 B CN 113605884B
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- 238000000034 method Methods 0.000 title claims abstract description 28
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- 238000009826 distribution Methods 0.000 description 8
- 230000005540 biological transmission Effects 0.000 description 5
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- E—FIXED CONSTRUCTIONS
- E21—EARTH OR ROCK DRILLING; MINING
- E21B—EARTH OR ROCK DRILLING; OBTAINING OIL, GAS, WATER, SOLUBLE OR MELTABLE MATERIALS OR A SLURRY OF MINERALS FROM WELLS
- E21B47/00—Survey of boreholes or wells
- E21B47/12—Means for transmitting measuring-signals or control signals from the well to the surface, or from the surface to the well, e.g. for logging while drilling
- E21B47/14—Means for transmitting measuring-signals or control signals from the well to the surface, or from the surface to the well, e.g. for logging while drilling using acoustic waves
- E21B47/18—Means for transmitting measuring-signals or control signals from the well to the surface, or from the surface to the well, e.g. for logging while drilling using acoustic waves through the well fluid, e.g. mud pressure pulse telemetry
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- E—FIXED CONSTRUCTIONS
- E21—EARTH OR ROCK DRILLING; MINING
- E21B—EARTH OR ROCK DRILLING; OBTAINING OIL, GAS, WATER, SOLUBLE OR MELTABLE MATERIALS OR A SLURRY OF MINERALS FROM WELLS
- E21B47/00—Survey of boreholes or wells
- E21B47/12—Means for transmitting measuring-signals or control signals from the well to the surface, or from the surface to the well, e.g. for logging while drilling
- E21B47/14—Means for transmitting measuring-signals or control signals from the well to the surface, or from the surface to the well, e.g. for logging while drilling using acoustic waves
- E21B47/18—Means for transmitting measuring-signals or control signals from the well to the surface, or from the surface to the well, e.g. for logging while drilling using acoustic waves through the well fluid, e.g. mud pressure pulse telemetry
- E21B47/20—Means for transmitting measuring-signals or control signals from the well to the surface, or from the surface to the well, e.g. for logging while drilling using acoustic waves through the well fluid, e.g. mud pressure pulse telemetry by modulation of mud waves, e.g. by continuous modulation
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Abstract
The swing valve pulse generator comprises a motor, a stator, a rotor and a torsion shaft, wherein the motor drives the torsion shaft, and the torsion shaft drives the rotor to rotate relative to the stator to generate a pulse signal; the method comprises the following steps: determining a maximum closing angle of the rotor when the swing valve pulse generator is not powered on; determining a rigidity coefficient of the torsion shaft and an assembly angle of the torsion shaft according to preset optimization conditions of torsion shaft parameters; wherein the optimization conditions include: and the balance torque of the rotor in the maximum closing angle is less than or equal to 0, and the variation of the balance torque of the rotor in the swinging angle range is minimum. The embodiment of the disclosure can reasonably determine parameters of the swing valve pulse generator such as the assembly angle, the rigidity coefficient and the like of the torsion shaft, can avoid pump holding, and enables the swing valve pulse generator to have good performance.
Description
Technical Field
The present disclosure relates to a swing valve pulser, and more particularly, to a method of determining swing valve pulser torque axis parameters.
Background
In the petroleum and natural gas drilling process, the measurement while drilling system plays an irreplaceable role in improving drilling and production efficiency and shortening engineering period. The measurement while drilling system mainly comprises various downhole parameter measuring instruments, an information transmission tool and a ground receiving tool, wherein a drilling fluid pulse method represented by continuous wave signal transmission in the information transmission system is widely paid attention to at home and abroad.
A continuous wave pulse generator mechanism for swinging valve is composed of motor, torsion axle, stator and rotor. The torque shaft is driven by the direct current motor, and the rotor is driven by the torque shaft to rotate to generate a pulse signal. The torsion shaft is used as a connecting piece of the motor and the rotor and a power transmission device. However, in determining parameters of the swing valve pulser (such as the assembly angle and stiffness coefficient of the torsion bar), the parameters are selected by experience, which is disadvantageous in improving the performance of the swing valve pulser.
Disclosure of Invention
The following is a summary of the subject matter described in detail herein. This summary is not intended to limit the scope of the claims.
The embodiments of the present disclosure provide the following solutions.
The swing valve pulse generator comprises a motor, a stator, a rotor and a torsion shaft, wherein the motor drives the torsion shaft, and the torsion shaft drives the rotor to rotate relative to the stator to generate a pulse signal; the method comprises the following steps:
determining a maximum closing angle of the rotor when the swing valve pulse generator is not powered on;
determining a rigidity coefficient of the torsion shaft and an assembly angle of the torsion shaft according to preset optimization conditions of torsion shaft parameters;
wherein the optimization conditions include: and the balance torque of the rotor in the maximum closing angle is less than or equal to 0, and the variation of the balance torque of the rotor in the swinging angle range is minimum.
The embodiment of the disclosure can reasonably determine parameters of the swing valve pulse generator such as the assembly angle, the rigidity coefficient and the like of the torsion shaft, can avoid pump holding, and enables the swing valve pulse generator to have good performance.
Other aspects will become apparent upon reading and understanding the accompanying drawings and detailed description.
Drawings
The accompanying drawings are included to provide a further understanding of the present technology and are incorporated in and constitute a part of this specification, illustrate the technology and together with the examples of the application, and do not constitute a limitation to the technology.
FIG. 1 is a schematic diagram of a swing valve pulser;
FIG. 2 is a schematic cross-sectional view of a wobble valve pulser stator and rotor;
FIG. 3 is a graph of hydraulic torque distribution along rotor angle for an embodiment of the present application;
FIG. 4 shows a displacement of 2.0m according to an embodiment of the present application 3 A distribution curve graph of the balance torque along the rotor angle at different rotor assembly angles at/min;
FIG. 5 is a graph showing the distribution of balancing torque along rotor angle at different displacements for an embodiment of the present application with a rotor assembly angle of 8.19;
FIG. 6 is a graph of the distribution of balancing torque along rotor angle after torque parameters are optimized for different displacements;
FIG. 7 is a displacement of 2.0m 3 A distribution curve graph of ideal torsion shaft torque along the rotor angle at/min;
FIG. 8 is a displacement of 1.8m 3 A distribution curve graph of ideal torsion shaft torque along the rotor angle at/min;
FIG. 9 is a graph of the equilibrium torque at different torsional stiffness coefficients;
FIG. 10 is a graph of the balancing torque at different torsion bar assembly angles;
FIG. 11 is a low displacement balancing torque after optimizing torque parameters based on high displacement;
FIG. 12 is a high displacement balancing torque after optimizing torque parameters based on low displacement;
FIG. 13 is a linear relationship of torque shaft maximum stiffness coefficient to motor torque rating;
FIG. 14 is a raw and smoothed rotor angle control trajectory (40 Hz);
FIG. 15 is a raw and smoothed rotor speed profile (40 Hz);
FIG. 16 is a raw and smoothed rotor drive power profile (40 Hz);
FIG. 17 is rotor power and balancing torque (40 Hz);
fig. 18 shows rotor power and rotational speed (40 Hz).
Detailed Description
The following description of the embodiments of the present application will be made clearly and fully with reference to the accompanying drawings, in which it is evident that the embodiments described are only some, but not all, of the embodiments of the present application. All other embodiments, which can be made by one of ordinary skill in the art based on the embodiments herein without making any inventive effort, are intended to be within the scope of the present application.
It should be noted that all directional indicators (such as up, down, left, right, front, and rear) in the embodiments of the present application are merely used to explain the relative positional relationship, movement conditions, and the like between the components in a specific posture (as shown in the drawings), and if the specific posture is changed, the directional indicators are correspondingly changed.
In addition, descriptions such as those related to "first," "second," and the like in the embodiments of the present application are for descriptive purposes only and are not to be construed as indicating or implying a relative importance or implying a number of technical features which are indicated. Thus, a feature defining "a first" or "a second" may explicitly or implicitly include at least one such feature. In the description of the present application, the meaning of "plurality" is at least two, such as two, three, etc., unless explicitly defined otherwise.
In the present application, unless explicitly specified and limited otherwise, the terms "coupled," "secured," and the like are to be construed broadly, and for example, "secured" may be either permanently attached or removably attached, or integrally formed; can be mechanically or electrically connected; either directly or indirectly, through intermediaries, or both, may be in communication with each other or in interaction with each other, unless expressly defined otherwise. The specific meaning of the terms in this application will be understood by those of ordinary skill in the art as the case may be.
In addition, the technical solutions of the embodiments of the present application may be combined with each other, but it is necessary to be based on the fact that those skilled in the art can implement the technical solutions, and when the technical solutions are contradictory or cannot be implemented, the combination of the technical solutions should be considered to be absent, and is not within the scope of protection claimed in the present application.
Fig. 1 is a schematic diagram of a swing valve pulser according to an embodiment of the present disclosure. As shown in fig. 1, the swing valve pulse generator provided by the embodiment of the disclosure includes a motor 2, a torsion shaft 3, a coupler 4, a stator 5 and a rotor 6 mounted on a bearing 1, the torsion shaft 3 is driven by the motor 2, and the torsion shaft 3 drives the rotor 6 to rotate through the coupler 4 to generate a pulse signal. Mud flows from the side where the stator 5 is located to the side where the rotor 6 is located, the motor 2 drives the torsion shaft 3, the torsion shaft 3 drives the rotor 6 to rotate relative to the stator 5, the stator 5 and the rotor 6 of the swing valve pulse generator periodically block a fluid channel, and the mud passing through is sheared, so that a pulse signal (also called pressure wave) is generated. The torsion shaft 3 can be realized by a torsion spring, for example.
The flow channel model of the slurry passing through the swing valve pulse generator adopts unstructured grids to carry out dispersion, and as the rotor 6 carries out reciprocating motion, the grid updating is realized by adopting a dynamic grid technology, and the flow field characteristics of the rotor 6 of the swing valve pulse generator under different rotating speed conditions are simulated by Fluent software. Under the condition that the flow field is stable, the momentum change of the slurry generates steady-state hydraulic torque, so that the flow field can be simulated at different angles by controlling the rotor 6 to move at a uniform speed or to be stationary, and the steady-state hydraulic torque characteristic can be analyzed. The dynamic hydraulic torque is caused by flow field change caused by acceleration and deceleration of the rotor 6, and the dynamic hydraulic torque characteristic of the rotor 6 can be analyzed by controlling the rotor 6 to move under a certain acceleration. By fluid analysis of the stator and rotor calculation domains of the swing valve pulser, the following conclusions can be drawn: the hydraulic torque is distributed along the same direction, namely, the rotor 6 is driven to move towards the closing direction; the hydraulic torque has close relation with the position of the rotor 6, and has smaller relation with the angular velocity and the angular acceleration of the rotation of the rotor 6; pressure wave changes are primarily related to flow, i.e., valve port flow area; the influence of the rotation angle of the rotor 6 on the hydraulic torque is obvious, and the maximum hydraulic torque appears when the stator and the rotor are close to the closed position. According to theoretical analysis, experimental tests can be combined again, and a relation curve of hydraulic torque (steady hydraulic torque and dynamic hydraulic torque) and the rotation angle of the rotor can be obtained. Under the same condition, the change rule of the dynamic hydraulic torque is basically the same as the steady hydraulic torque, and the amplitude is slightly lower than the steady hydraulic torque. Therefore, the hydraulic torque to which the rotor 6 of the present embodiment is subjected may be calculated as the steady-state hydraulic torque, and the margin of the calculation result may be increased while being calculated as the steady-state hydraulic torque, but may be calculated as the dynamic hydraulic torque in other embodiments, or may be calculated as a weighted value of the steady-state hydraulic torque and the dynamic hydraulic torque.
The inventor finds that the rigidity and the assembly angle of the torsion shaft 3 in the swing valve pulse generator directly influence the working effect of the pulse generator, and the elastic deformation torque generated by the torsion shaft 3 when the swing valve pulse generator works can help the motor to overcome the external hydraulic torque, but if the restoring torque generated by deformation of the torsion shaft 3 is overlarge, the load of the motor can be increased, and the effect of a mud pulse signal is influenced. For example, if a smaller angle assembly is selected, there may be problems such as excessive power consumption of the motor, slow response speed, and the like. If a larger angle is selected for assembly, if the hydraulic torque at the angle is greater than the torsion shaft torque, the rotor 6 tends to close under the action of the hydraulic torque when the instrument is not powered on or fails, thereby causing pump holding. It is therefore desirable to provide a method for computationally checking the torsion shaft 3 such that the torque of the torsion shaft 3, the hydraulic torque of the mud, and the function of the motor are matched.
As shown in fig. 2, a schematic cross-sectional view of the structure of the wobble valve pulser stator 5 and rotor 6 according to an exemplary embodiment of the present application is shown. In the illustrated example, the blades of the stator 5 and the rotor 6 are sector surfaces, the angles are alpha, the opening angles are gamma, and the torsion shaft assembly angle is beta 0 I.e. the initial angle between the installed stator and rotor blades: beta is not less than 0 DEG 0 And gamma is less than or equal to gamma. In an example, the sector angles of the stator blades and the rotor blades are each 37 °, the opening angles are each 23 °, and hereinafter, 23 ° is taken as an example of the opening angles, but the present application is not limited thereto.
In order to enable the torsion shaft 3 to balance the hydraulic torque, the torsion shaft assembly angle β 0 Typically greater than 0. The oscillating valve pulse generator can not work normally when the pump is started, if the torsion shaft 3 is improperly assembledThe torque generated by the shaft 3 is not able to balance the hydraulic torque, which may drive the rotor 6 closed, thus causing a pump hold. Assuming that the rigidity coefficient of the torsion shaft 3 is K, when the angle of the rotor 6 is β, the hydraulic torque is T H Beta, torque of torsion shaft is T M (β)＝K·(β 0 - β), the balance torque of which is T (β): then there are:
T(β)＝T H (β)+T M (β)＝T H (β)+K·(β 0 -β)
when the balancing torque T (β) >0, the balancing torque indicates that the rotor 6 is driven to rotate in the closing direction of the swing valve pulser.
The application provides a method for determining a torsion shaft assembly angle of a swing valve pulse generator, which comprises the following steps:
determining the maximum closing angle of the rotor when the swing valve pulse generator is not electrified;
and determining the assembly angle of the torsion shaft, wherein when the rotor is at the maximum closing angle, the torque generated by the torsion shaft is opposite to the hydraulic torque received by the rotor, and the value of the torque is not smaller than the value of the hydraulic torque.
In an exemplary embodiment of the present disclosure, after the design and the selection of the swing valve pulse generator are completed, by properly selecting the torsion shaft assembly angle, a certain opening degree of the rotor 6 is ensured before the swing valve pulse generator is powered on or in the case of failure, so that pump holding can be avoided. Assume that the maximum closing angle of the rotor 6 (the maximum angle of rotation from the fully open position of the rotor 6 to the closing direction) when the swing valve pulser is not powered on in the operating scene is θ C (θ C < 23 °), the present embodiment defines that the maximum closing angle is the maximum value of the closing angle of the rotor 6 when the swing valve pulser can maintain the normal displacement cycle, and the corresponding steady-state hydraulic torque is denoted as T H (θ C ) The conditions for avoiding pump holding in this example are:
T(θ C )＝K·(β 0 -θ C )+T H (θ C )≤0
the method can be as follows:
balance torque T (θ) C ) Less than or equal to 0 means that the rotor angle is θ C When the rotor 6 is not driven to rotate in the closing direction under the combined action of the steady hydraulic torque and the torsion shaft torque, the rotor 6 is prevented from rotating to be more than theta C The angle of the valve pulse generator makes it difficult to maintain normal displacement, and pump is suppressed. That is, the maximum closing angle of the rotor 6 when not energized is θ C In the time of, for the torsion shaft 3 with the rigidity coefficient of K, the torsion shaft assembly angle beta of the pump can be prevented 0 Can not exceed the maximum value of
In one example, assuming that the rotor 6 is controlled in a segmented sinusoidal manner, i.e. a certain voltage is applied to the motor windings, a segmented sinusoidal current is generated in the motor windings, and the purpose of controlling the motor torque is achieved by controlling the amplitude and phase of the sinusoidal current. When the swing valve pulse generator is operating normally (i.e. the displacement is normal), its equivalent flow area is about 33% of the full open rotor 6, i.e. the maximum closing angle of the rotor 6 is set to 67% of the opening angle of the vanes of the stator 5, corresponding to a rotor angle of 23 ° x 0.67=15.41°, so that when the rotor is at its maximum closing angle θ C Pulser can maintain normal displacement cycle when=15.41°, θ determined by this example C Bringing into the solution of the torsion shaft assembly angle beta 0 The torque assembly angle at this time can be obtained by the formula (1):
further determination of the hydraulic torque at 15.41 ° of rotor 6 and the stiffness coefficient of torsion shaft 3 allows determination of a specific torsion shaft assembly angle β 0 。
The torsion shaft assembly angle beta 0 Maximum assembly angle for the torque shaft 3. I.e. if the torque shaft assembly angle exceeds beta 0 When the motor is not powered on or after the swing valve pulse generator fails, the torque of the torsion shaft cannot completely resist the steady-state hydraulic torque, and the swing valve pulse generator cannot maintain the normal displacement cycle. However, if the fitting angle β of the torsion shaft 3 0 Too small, too large restoring moment generated by deformation of the torsion shaft 3 can increase the load of the motor and affect the effect of the mud pulse signal. Therefore, on the premise of preventing the pump from being blocked, the torsion shaft assembly angle beta 0 Nor too small. Thus, the torsion shaft assembly angle beta is determined 0 After the upper limit of (2), the lower limit of the value needs to be further defined. In an exemplary embodiment of the present disclosure, the angle of assembly of the torsion shaftWherein A is a set coefficient. Further, the range of the coefficient A is generally set to be 0.7.ltoreq.A.ltoreq.0.9 in consideration of engineering requirements in terms of processing, assembly, reliability and the like.
FIG. 3 is a graph of two curves, each showing the mud displacement of 1.8m, respectively, in an exemplary embodiment of the present disclosure 3 /min and 2.0m 3 The rotor 6 is subjected to hydraulic torque at different angles under the condition of/min. The gap between the stator 5 and the rotor 6 of this embodiment is 1.26mm. Taking the torsion rigidity coefficient k=10/24 n·m/° as an example, when the rotor 6 is at the maximum closing angle of 15.41 °, the hydraulic torque is 2.3635n·m (the mud displacement is 1.8 m), respectively 3 Per min) and 3.0061 N.m (mud displacement of 2.0 m) 3 At/min), the rotor assembly angle beta of the swing valve pulse generator can be obtained by the formula for calculating the torsion shaft assembly angle 0 9.7377 ° and 8.1953 °, respectively. In other words, in the present embodiment, when the torsion shaft assembly angle does not exceed 9.7377 °, the displacements of 1.8m respectively can be ensured even if the instrument is powered off 3 The pump is not blocked in the time of/min. When the torsion shaft assembly angle is not more than 8.1953 DEG, the displacement can be ensured to be 2.0m even if the instrument is powered off 3 The pump is not blocked in the time of/min.
At a mud displacement of 2.0m 3 The calculation result at/min is taken as an example, when the number is increasedThe balance torque of the hydraulic torque and the torque of the torsion shaft acting on the rotor 6 at the time of large or small rotor assembly angle is shown in fig. 4. As can be seen from FIG. 4, when beta 0 About 8.19 DEG at 2.0m 3 In the case of a displacement outage/min, the rotor 6 has a balancing torque of 0 at 10 ° to 16 °, the rotor 6 being substantially in oscillation between 10 ° and 16 °. At this time, when the rotor 6 is controlled to swing around the angle range by the motor to shear the slurry to generate a pulse signal, the balance torque to be overcome is small, so that the power consumption is low and the response speed is high.
If beta is 0 When the rotor 6 is turned off by about 10 °, the equilibrium torque becomes 0, and the opening angle of the rotor 6 (the angle from the rotor current position to the fully closed position) can be maintained only by about 3 °. If beta is 0 When the closing angle of the rotor 6 is reduced to 7 °, the equilibrium torque becomes 0 when the closing angle is around 5 °, and the opening angle of the rotor 6 becomes large, so that the rotor can be maintained at about 18 °. It can be seen that beta is reduced 0 The opening angle of the rotor 6 is increased, the stator 5 and the rotor 6 can provide larger overflow area for slurry to pass through, but balance torque to be overcome when the rotor 6 is at a higher angle is increased (because the torque is increased), so that the power consumption is increased, and the response speed is reduced; conversely, if beta is increased 0 The opening angle of the rotor 6 is reduced, so that the flow area of the slurry is reduced, and the pump is possibly blocked.
Likewise, the mud displacement amounts are respectively 1.8 and 2.0m 3 The calculation of/min is exemplified when a larger displacement (2.0 m in the example is used 3 After calculating the rotor assembly angle, the rotor assembly angle is directly applied to the smaller displacement condition (1.8 m in the example) 3 In/min), as shown in FIG. 5, when the rotor is assembled at an angle beta 0 About 8.19 deg., if the displacement is reduced to 1.8m 3 The opening angle of the rotor 6 is increased to a certain extent and can be kept at about 14 degrees, and the maximum closing angle of the rotor 6 is about 9 degrees at the moment; conversely, if the displacement increases beyond 2.0m 3 The opening angle of the rotor 6 is reduced, i.e. the maximum closing angle of the rotor 6 exceeds the angle at which the normal cycle of displacement can be maintained, i.eAbove 15.41 °, the flow area of the slurry becomes smaller in this case, possibly resulting in pump holding.
If the swing valve pulser has multiple displacement conditions, the hydraulic torque may be calculated from the hydraulic torque experienced by the swing valve pulser when at the highest displacement condition of the multiple displacement. If the swing valve pulse generator has three working conditions of high displacement, medium displacement and low displacement, the hydraulic torque can be calculated according to the hydraulic torque received by the swing valve pulse generator when the swing valve pulse generator is in the medium displacement working condition.
In an exemplary embodiment of the present disclosure, the mud displacement is divided into 1.8 and 2.0m 3 Two different conditions per min. When using 2.0m 3 When the rotor assembly angle is calculated by the per-min displacement, the initial slurry displacement is 1.8 or 2.0m no matter the pump is started 3 Every min, can prevent the pump from being blocked; when the instrument fails, the mud discharge capacity is not required to be 2.0m 3 The/min is reduced to 1.8m 3 The maximum closing angle of the rotor 6 can also maintain the angle of normal circulation of the displacement.
When 1.8m is used 3 When the rotor assembly angle is calculated by the per-min displacement, the swing valve pulse generator can be normally opened and closed under the condition that the motor 2 works normally, but the motor 2 is not normally electrified at the initial stage of pump opening, and if the slurry displacement is set to be 2.0m at the moment 3 In order to avoid pump hold-down at the initial stage of pump start, the initial stage of slurry discharge should be set to 1.8m 3 Displacement per min; similarly, when the instrument fails, the displacement should be reduced to 1.8m 3 And/min to prevent pump holding.
The method for determining the torsion shaft assembly angle of the swing valve pulse generator can reasonably determine the torsion shaft assembly angle, and can reduce the power consumption of the swing valve pulse generator as much as possible and improve the response speed of the swing valve pulse generator on the premise of avoiding pump holding.
The application also provides a swing valve pulse generator, and this swing valve pulse generator includes motor, stator, rotor and torsion bar, and the motor passes through the torsion bar and drives the rotor and rotate for the stator. The assembly angle of the torsion shaft is determined according to the method for determining the assembly angle of the torsion shaft of the swing valve pulse generator.
In an exemplary embodiment of the present application, after the stator and rotor structures (including the clearances) are determined, the hydraulic torque curves thereof are determined at different displacements. Before the rigidity coefficient of the torsion shaft 3 is not determined, the requirements of hydraulic torque balance and pump suppression avoidance are comprehensively considered, and besides the torsion shaft assembly angle for pump suppression avoidance is calculated, the rigidity coefficient of the torsion shaft 3 is determined. Considering that the dynamic hydraulic torque and the steady-state hydraulic torque are comparable, the rigidity factor of the torsion shaft 3 is determined based on the steady-state hydraulic torque.
The variation of the balancing torque to which the rotor 6 is subjected during the oscillation is a minimum of optimization conditions. Based on the distribution characteristics of the hydraulic torque along the rotor angle (as shown in fig. 3) and the linear characteristics of the torque shaft torque, if the optimization condition is satisfied, the balance torque monotonically decreases over the variation range of the rotor angle β (as shown in fig. 4). The balancing torque T (β) has an extreme value when the rotor angle β is 0 ° and 23 °, and the balancing torque directions are opposite. If |t (0) |= |t (23) |, the magnitude of T (β) is minimum.
Meanwhile, considering the condition that the balance torque needs to be met when the pump is not held down, taking the opening angle of the stator and the rotor as 23 DEG as an example, the following equation set can be obtained after the combination:
assuming that the minimum value of the rigidity coefficient of the torsion shaft 3 is K 0 Assembly angle beta of rotor 6 0 Is beta at maximum value 0-max Then:
thereby obtaining
Wherein θ C For maximum closing angle (maximum angle of rotation from fully open position of rotor 6 to closing direction) of rotor 6 when swing valve pulser is not powered on in working scene, θ C < 23 deg.. The present embodiment defines that the maximum closing angle is the maximum value of the closing angle of the rotor 6 when the swing valve pulser is able to maintain a normal displacement cycle.
The rotor 6 is supposed to be controlled in a sectional sine mode, namely, a certain voltage is applied to the motor winding, so that sectional sine current is generated in the motor winding, and the purpose of controlling the motor torque is achieved by controlling the amplitude and the phase of the sine current. When the swing valve pulse generator is operating normally (i.e. the displacement is normal), its equivalent flow area is about 33% of the full open rotor 6, i.e. the maximum closing angle of the rotor 6 is set to 67% of the opening angle of the vanes of the stator 5, corresponding to a rotor angle of 23 ° x 0.67=15.41°, so that when the rotor is at its maximum closing angle θ C The pulser is able to maintain a normal displacement cycle when =15.41°.
In one illustrative embodiment, based on 1.8 and 2.0m 3 Hydraulic torque of/min displacement, respectively calculating corresponding K 0 And beta 0 The result is:
as can be seen from fig. 6, the balance torque at both displacements is substantially the same by optimization of the method described above. The balancing torque is distributed along the rotor angle in a monotonically decreasing manner. In the range of rotor angle, the balance torque of the middle section is approximately 0. At the two end points of the rotor angle value range (namely 0 DEG and 23 DEG), the balance torque has extreme values with equal magnitude and opposite directions. At this time, the rotor 6 undergoes minimal change in the balance torque during the swinging, and the result exhibits an optimization effect.
FIGS. 7 and 8 show displacements of 2.0 and 1.8m, respectively 3 Hydraulic torque T at/min H (beta) relation to the desired torsion axis torque. T (T) H And (beta) is basically linear when the rotor angle is 9-17 degrees, and the slope of the section of curve is the ideal rigidity coefficient of the torsion shaft 3. T (T) H The rotor angle when the value of (beta) is 0 is the torsion shaft assembly angle, namely T H (β 0 )＝0。
Beta retention 0 Unchanged, change K 0 The effect of (2) is shown in fig. 9. When the torsional rigidity coefficient K is increased 0 Time (k=k in fig. 9) 0 +0.05) will result in an increase in the balance torque of the rotor angle beta in both the 0 deg. and 23 deg. states, thereby increasing the demand for motor output torque. When the torsional rigidity coefficient K is reduced 0 Time (k=k in fig. 9) 0 0.05), the rotor angle beta decreases in the balance torque in both the 0 deg. and 23 deg. states, but in the linear phase of the hydraulic torque, the torsion shaft 3 does not balance the hydraulic torque well in this phase. Also, hold beta 0 Unchanged, reduce K 0 The equilibrium torque curve rotates counterclockwise along point (8.18 °, 0) by a certain angle. The rotation of the curve makes the balance torque greater than 0 when the rotor 6 is at the maximum closing angle, i.e. the balance torque drives the rotor 6 to be continuously closed, so that the swing valve pulse generator cannot maintain the normal displacement cycle, and thus pump holding may be caused.
As shown in fig. 10, hold K 0 Unchanged, but only change beta 0 . Since the stiffness coefficient of the torsion shaft 3 is not changed, only the intercept of the torsion shaft 3 on the Y-axis is changed (refer to fig. 7 and 8, but the torsion shaft torque curve after the torsion shaft assembly angle is changed is not shown in the drawings), the stiffness coefficient of the torsion shaft 3 is still matched with the slope of the hydraulic torque in the linear section, but the angle is relative to beta 0 The case of=8.18° is offset to some extent. This is shown in fig. 10 as the equilibrium torque curve translates approximately up and down the Y-axis.
If the torsion axis parameter (K 0 ，β 0 ) Based on high displacement 2.0m 3 Optimizing/min, and directly applying the torsion shaft parameter to low-displacement 1.8m 3 The distribution of the balancing torque along the rotor angle at the operating condition/min is shown in fig. 11. Otherwise, thenAs shown in fig. 12. As can be seen from fig. 11, the torque parameters calculated by the high displacement condition are directly applied to the low displacement condition, and when the swing valve pulser is not powered on or fails, the rotor 6 can obtain a larger opening angle. However, at the same time, the balancing torque applied to the rotor 6 in the completely closed state (the rotor angle is 23 °) is increased, and thus a higher demand is placed on the output torque of the motor. In contrast, as shown in fig. 12, the torque parameter calculated by the low displacement condition is directly applied to the high displacement condition, and the balance torque to which the rotor 6 is subjected in the fully closed state (rotor angle of 23 °) is reduced. However, under this torque parameter, the torque provided by the torque shaft 3 does not completely cancel the hydraulic torque, i.e. the balancing torque of the two still drives the rotor 6 closed. Therefore, in this case, if the swing valve pulser is not powered on at the initial stage of the start-up or if the swing valve pulser fails to power off in the middle, a pump-holding phenomenon may occur. Thus, considering both downhole safety and the requirement of reducing the balance torque, and considering the effect of mud density on hydraulic torque, the torque parameters should be determined with medium stator and rotor clearances, with clear water, medium (or medium offset) displacement.
Given the displacement of slurry, comprehensively considering the requirements of hydraulic torque balance and pump holding avoidance, and obtaining K according to the method 0 Is the minimum torsion rigidity coefficient for guaranteeing the safety in the pit. Considering the machining precision error of the torsion shaft 3 and the error generated by the swing valve pulse generator in the actual manufacturing and assembling process (the torsion shaft assembling angle is fixed), the maximum value K of the rigidity coefficient needs to be determined MAX Thereby obtaining the allowable range (K 0 ,K MAX )。
Assuming that the maximum output torque of the motor is MT, the steady-state hydraulic torque under the minimum displacement is T HL Beta, when K is greater than or equal to K 0 Balance torque |T (β) |+|T (23) |, therefore:
T(23)＝T HL (23)+K·(β 0 -23)≥-MT
thus, in theory, when the torsional rigidity coefficient satisfies K 0 ≤K≤K MAX By adopting the motor with rated output torque of MT, the rotor control within the specified displacement range can be realized. The rated output torque MT is related to the motor speed.
For example, torsion axis parameter (K 0 ,β 0 ) Based on 2.0m 3 The/min displacement determination, assuming a minimum displacement of 1.8m 3 /min, calculated K MAX The relationship with MT is shown in fig. 13. If the maximum output torque mt=5n·m of the motor, K 0 ＝0.4158≤K≤0.5。
An exemplary embodiment of the present application also provides a method for determining a motor parameter.
The rotation angle and speed of the rotor are determined by the modulation scheme and the transmission rate. The rotation speed of the motor is determined by the rotation angle of the rotor and the transmission rate (the movement track of the rotor). The motor speed and output torque should be capable of meeting rotor swing requirements. The motor output power can meet the requirements of the rotation angle and the speed of the rotor under the condition of resisting resistance caused by factors such as the torsional rigidity coefficient, the assembly angle and the like.
The rotation angle of the rotor 3 is controlled in a piecewise sinusoidal manner with a maximum frequency of 40Hz and fig. 14 is a rotation angle trace (a=2.0, b=0.86) of the rotor at 40 Hz. As shown in fig. 15, the minimum requirements for motor speed before and after track smoothing are about 1000rpm and 650rpm, respectively.
The maximum output torque MT of the motor is determined by determining the maximum value K of the torsional rigidity coefficient MAX Is the inverse of (1), i.e. given the torque axis parameter (K 0 ,β 0 ) And stator and rotor clearances, when the displacement is minimum and the rotor angle is 23 DEG, the balance torque has the maximum amplitude T MAX Therefore, it is theoretically required that the stationary torque MT. Gtoreq.T of the motor MAX . Fig. 13 also shows the relationship between the motor minimum torque and the torsion rigidity coefficient.
Assuming that the dynamic hydraulic torque is the same as the steady-state hydraulic torque, the power of the motor driving the rotor to move is as follows without considering the influence of moment of inertia:
wherein,to balance the torque (N.m),. About.>ω (t, f) is rotor rotational speed (r/s) and P (t, f) is in W, which is the rotor trajectory (°). When f=40 Hz, 1.8m based 3 The peak power of the original trajectory is 105W and the peak power of the smoothed trajectory is 84W, as shown in fig. 16.
Fig. 17 and 18 also show the profile of the rotor power versus the balancing torque at 40Hz and the profile of the rotor power versus the rotational speed at 40Hz, respectively.
Compared with the prior art, the method for determining the torsion shaft parameters of the swing valve pulse generator can calculate the torsion shaft assembly angle and the torsion shaft rigidity coefficient, and when the swing valve pulse generator is powered on or fails, the swing valve pulse generator can maintain normal displacement circulation when the rotor is at the maximum closing angle of the rotor when the rotor is not powered on. Meanwhile, the torque shaft assembly angle and the torque shaft rigidity coefficient calculated by the method can minimize the change of the balance torque born by the rotor in the swinging process on the premise of avoiding the pump holding, so that the load of the motor is reduced, the power of the motor is saved, the requirement on the motor parameter is reduced, and the response speed of the pulse signal is improved.
While the embodiments disclosed in the present disclosure are described above, the embodiments are only employed for facilitating understanding of the present disclosure, and are not intended to limit the present disclosure. Any person skilled in the art to which this disclosure pertains will appreciate that numerous modifications and changes in form and details of implementation can be made without departing from the spirit and scope of the disclosure, but the scope of the disclosure is to be determined by the appended claims.
Claims (8)
1. The swing valve pulse generator comprises a motor, a stator, a rotor and a torsion shaft, wherein the motor drives the torsion shaft, and the torsion shaft drives the rotor to rotate relative to the stator to generate a pulse signal; the method comprises the following steps:
determining a maximum closing angle of the rotor when the swing valve pulse generator is not powered on, wherein the maximum closing angle of the rotor is set to be the maximum value of the closing angle of the rotor when the swing valve pulse generator can maintain a normal displacement cycle;
determining a rigidity coefficient of the torsion shaft and an assembly angle of the torsion shaft according to preset optimization conditions of torsion shaft parameters;
wherein the optimization conditions include: the balance torque of the rotor is smaller than or equal to 0 when the rotor is at the maximum closing angle, and the change of the balance torque of the rotor in the swinging angle range is minimum; the balance torque is the sum of hydraulic torque and torque of the torsion shaft;
the optimization conditions are set as follows:
wherein T (beta) is the balance torque when the rotation angle of the rotor is beta, and beta is more than or equal to 0 and less than or equal to beta max ,β max Is the opening angle of the rotor, theta C T is the maximum closing angle of the rotor H And (beta) is the hydraulic torque when the rotation angle of the rotor is beta, K (beta) 0 -beta) is the torque of the torsion shaft at a rotation angle beta of the rotor, beta 0 The assembly angle of the rotor is that K is the rigidity coefficient of the torsion shaft;
determining the stiffness coefficient of the torsion shaft and the assembly angle of the torsion shaft comprises:
from the solution of formula (1) and formula (2) when T (β) =0:
wherein K is 0 Is the minimum value of the rigidity coefficient of the torsion shaft, beta 0-max For the angle beta of assembly of the rotor 0 Maximum value of T H (θ C )、T H (0) And T H (β max ) The value of the valve is determined according to the corresponding relation between the hydraulic torque of the swing valve pulse generator under the set displacement and the rotation angle of the rotor.
2. The method of claim 1, wherein: the rotor is controlled in a segmented sinusoidal mode, and the maximum closing angle of the rotor is set to be 65% -70% of the opening angle of the blades of the stator.
3. The method of claim 1, wherein:
determining the assembly angle of the torsion shaft further comprises: setting the assembly angle beta of the torsion shaft 0 Is the minimum value of (2):a is a set coefficient, and A is more than or equal to 0.7 and less than or equal to 0.9.
4. A method according to claim 1 or 3, wherein:
determining a stiffness coefficient of the torsion shaft, further comprising: determining the maximum value of the rigidity coefficient of the torsion shaft according to the following formula:
wherein K is MAX At maximum value of rigidity coefficient of torsion shaft, MT max T is the maximum value of the output torque of the motor HL (β max ) The swing valve pulse generator works at the minimum displacement and the rotation angle of the rotor is beta max When (1)Hydraulic torque.
5. The method of claim 4, wherein:
the method further comprises the steps of: the output torque MT of the motor is determined according to the following equation:
MT＝G(β max -η)-T HL (β max )
wherein G is according to K 0 And K MAX The rigidity coefficient of the selected torsion shaft is eta as beta 0-max The assembly angle of the torsion shaft is selected.
6. The method of claim 1, wherein:
the swing valve pulse generator has a plurality of displacement working conditions, and the hydraulic torque is calculated according to the hydraulic torque received by the swing valve pulse generator when the swing valve pulse generator is in the working condition of the highest displacement of the plurality of displacement working conditions.
7. The method of claim 1, wherein:
the swing valve pulse generator has three working conditions of high displacement, medium displacement and low displacement, and the hydraulic torque is calculated according to the hydraulic torque received by the swing valve pulse generator when the swing valve pulse generator is in the medium displacement working condition.
8. A pendulum valve pulser comprising a motor, a stator, a rotor and a torsion shaft, said motor driving said rotor to rotate relative to said stator via said torsion shaft, the parameters of said pendulum valve pulser being determined according to the method of any one of claims 1 to 7.
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| CN205067013U (en) * | 2015-10-21 | 2016-03-02 | 中国海洋石油总公司 | Rocker valve mud impulse generator torsional spring torsion -testing calibration device |
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| US20210189873A1 (en) * | 2019-12-18 | 2021-06-24 | Baker Hughes Oilfield Operations Llc | Oscillating shear valve for mud pulse telemetry and operation thereof |
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| CN205067013U (en) * | 2015-10-21 | 2016-03-02 | 中国海洋石油总公司 | Rocker valve mud impulse generator torsional spring torsion -testing calibration device |
| CN105604543A (en) * | 2015-12-18 | 2016-05-25 | 中国海洋石油总公司 | Rocking valve mud pulse generator transmission system |
| CN105785268A (en) * | 2016-04-07 | 2016-07-20 | 中国海洋石油总公司 | Calibration method for slurry pulse generator of shearing valve |
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