A method for controlling multiple hydraulic cylinders of a submarine clock door synchronously

Abstract

The application relates to the field of deep-sea operation, and discloses a multi-oil-cylinder synchronous control method for a diving bell door, which comprises the following steps: acquiring environmental water pressure data outside the diving bell and cabin water pressure data inside the diving bell; obtaining an internal-external pressure difference value; calculating a maximum displacement deviation value among multiple hydraulic oil cylinders based on displacement data; calculating a displacement deviation threshold value based on the internal-external pressure difference value; comparing the displacement data of the multiple hydraulic oil cylinders to determine a leading oil cylinder and a lagging oil cylinder among the multiple hydraulic oil cylinders; calculating a difference value between the maximum displacement deviation value and the displacement deviation threshold value; obtaining a flow compensation coefficient based on the difference value and the internal-external pressure difference value; and generating a first oil supply instruction for controlling a first proportional valve and / or generating a second oil supply instruction for controlling a second proportional valve based on the flow compensation coefficient. Through the technical scheme, the technical problem that a fixed preset value cannot adapt to a deep-sea variable working condition environment in the related art is solved.

Description

Technical Field

[0001] This application relates to the field of deep-sea operations, and in particular to a method for synchronous control of multiple hydraulic cylinders in a diving bell hatch. Background Technology

[0002] As a piece of equipment connecting the surface mother ship and the underwater working environment, the doors of a diving bell usually need to withstand high deep-sea water pressure. In order to overcome the high water pressure and the weight of the door itself, the doors of diving bells generally use multiple hydraulic cylinders to drive them to open or close in a coordinated manner.

[0003] In related technologies, multi-cylinder synchronous control uses sensors to acquire the extension and retraction displacements of two or more cylinders and calculates the displacement difference. When the displacement difference is greater than a preset value, the oil supply of the cylinder that extends faster is reduced and the oil supply of the cylinder that extends slower is increased until the displacement difference returns to within the preset value.

[0004] The inventors discovered that the pressure difference between the inside and outside of the hatch is inconsistent on both sides of the diving bell at different diving depths and during the inflation and deflation stages. In shallow water, the pressure difference is small and the hatch is subjected to light force, which can allow for a large displacement difference of the hydraulic cylinders without jamming. The fixed preset value limits the displacement difference to a small range, which restricts the opening and closing speed. In deep water, the pressure difference is large, and after the hatch is squeezed by lateral water pressure, the displacement difference between the hydraulic cylinders is more likely to cause jamming. Under this condition, the fixed preset value may not be able to provide synchronization accuracy that is adapted to the current pressure difference. Summary of the Invention

[0005] This application provides a multi-cylinder synchronous control method for a diving bell hatch, which at least partially solves the above-mentioned technical problems.

[0006] To achieve the above objectives, this application provides a multi-cylinder synchronous control method for a diving bell hatch, comprising: Acquire the ambient water pressure data outside the diving bell and the water pressure data inside the chamber; calculate the difference between the ambient water pressure data and the chamber water pressure data to obtain the internal and external pressure difference value; Obtain displacement data of multiple hydraulic cylinders that drive the hatch to open; calculate the maximum displacement deviation value between the multiple hydraulic cylinders based on the displacement data; The displacement deviation threshold is calculated based on the internal and external pressure difference value. The displacement deviation threshold represents the maximum allowable displacement difference that does not cause door jamming under the current pressure difference condition. The internal and external pressure difference value is inversely proportional to the displacement deviation threshold. When the maximum displacement deviation value is greater than the displacement deviation threshold, the displacement data of the plurality of hydraulic cylinders are compared to determine the leading cylinder and the lagging cylinder among the plurality of hydraulic cylinders; the proportional valve that supplies oil to the leading cylinder is determined as the first proportional valve, and the proportional valve that supplies oil to the lagging cylinder is determined as the second proportional valve. Calculate the difference between the maximum displacement deviation value and the displacement deviation threshold; based on the difference and the internal and external pressure difference value, obtain the flow compensation coefficient; Based on the flow compensation coefficient, a first oil supply command is generated to control the first proportional valve, and / or a second oil supply command is generated to control the second proportional valve; the first oil supply command is used to reduce the oil supply to the leading cylinder, and the second oil supply command is used to increase the oil supply to the lagging cylinder.

[0007] In this embodiment, the above technical solution tightens the synchronization control precision under deep-water high-pressure differential conditions to prevent door jamming and seal damage, and relaxes the synchronization control precision under shallow-water low-pressure differential conditions to increase door opening speed, thus solving the technical problem in related technologies that fixed preset values ​​cannot adapt to the changing working conditions in deep sea.

[0008] Other features and advantages of this application will be described in detail in the following detailed description section. Attached Figure Description

[0009] To more clearly illustrate the technical solutions in the embodiments of this application, the accompanying drawings used in the description of the embodiments will be briefly introduced below. Obviously, the drawings described below are only some embodiments of this application. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0010] Figure 1 This is a flowchart illustrating the steps of a multi-cylinder synchronous control method for a diving bell hatch provided in an exemplary embodiment of this application. Detailed Implementation

[0011] The technical solutions of the embodiments of this application will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only a part of the embodiments of this application, and not all of them. All other embodiments obtained by those skilled in the art based on the embodiments of this application without creative effort are within the protection scope of this application.

[0012] The multi-cylinder synchronous control method for the diving bell hatch provided in this application embodiment can be applied to the hatch drive system of a diving bell. The hatch drive system includes: an external pressure sensor mounted on the surface of the diving bell's outer shell for collecting external water pressure values ​​as ambient water pressure data; an internal pressure sensor mounted on the inner wall of the diving bell for collecting internal chamber pressure values ​​as chamber water pressure data; multiple hydraulic cylinders, respectively arranged around the hatch, for jointly driving the opening and closing of the hatch; multiple proportional valves, each for controlling the flow rate of hydraulic oil supplied to a corresponding hydraulic cylinder; multiple displacement sensors, each for acquiring the extension and retraction displacement of its corresponding hydraulic cylinder; and a control unit communicatively connected to the external pressure sensor, internal pressure sensor, multiple displacement sensors, and multiple proportional valves.

[0013] Reference Figure 1 A method for synchronous control of multiple hydraulic cylinders in a diving bell hatch, comprising: Step 101: Obtain the ambient water pressure data outside the diving bell and the water pressure data inside the chamber; calculate the difference between the ambient water pressure data and the chamber water pressure data to obtain the pressure difference between the inside and outside.

[0014] Specifically, the control unit collects external water pressure values ​​as ambient water pressure data through external pressure sensors installed on the surface of the diving bell's outer shell, and collects internal chamber pressure values ​​as chamber water pressure data through internal pressure sensors installed on the inner wall of the diving bell. The ambient water pressure data refers to the external seawater pressure value at the current water depth position of the diving bell, and the chamber water pressure data refers to the current air pressure value of the internal chamber of the diving bell. The control unit subtracts the ambient water pressure data from the chamber water pressure data to obtain the internal and external pressure difference value, which reflects the pressure difference between the two sides of the hatch.

[0015] Step 102: Obtain the displacement data of multiple hydraulic cylinders that drive the hatch to open; calculate the maximum displacement deviation value between the multiple hydraulic cylinders based on the displacement data.

[0016] Specifically, the control unit obtains the current extension and retraction displacement value of each hydraulic cylinder through displacement sensors corresponding to each hydraulic cylinder; if there are N hydraulic cylinders in total, the maximum displacement deviation value is taken as the difference between the maximum and minimum extension and retraction displacement values ​​of each hydraulic cylinder, which is used to measure the degree of inconsistency in the current synchronous operation of multiple cylinders.

[0017] Step 103: Calculate the displacement deviation threshold based on the internal and external pressure difference value. The displacement deviation threshold represents the maximum allowable displacement difference that does not cause door jamming under the current pressure difference condition. The internal and external pressure difference value is inversely proportional to the displacement deviation threshold.

[0018] Specifically, the control unit calculates the displacement deviation threshold based on the internal and external pressure difference. When the internal and external pressure difference is large, i.e., in deep water with high pressure difference, the hatch is subjected to stronger lateral compression and has a lower tolerance for asynchronous operation of multiple cylinders. Therefore, the displacement deviation threshold is set to a smaller value to tighten the accuracy requirements of synchronous control. When the internal and external pressure difference is small, i.e., in shallow water with low pressure difference, the hatch is subjected to less force. Therefore, the displacement deviation threshold is set to a larger value to relax the accuracy requirements of synchronous control. The internal and external pressure difference and the displacement deviation threshold maintain an inverse proportional relationship, i.e., the larger the internal and external pressure difference, the smaller the displacement deviation threshold.

[0019] Step 104: When the maximum displacement deviation value is greater than the displacement deviation threshold, compare the displacement data of the multiple hydraulic cylinders to determine the leading cylinder and the lagging cylinder among the multiple hydraulic cylinders, and determine the proportional valve supplying oil to the leading cylinder as the first proportional valve, and the proportional valve supplying oil to the lagging cylinder as the second proportional valve. Specifically, the control unit compares the maximum displacement deviation value with the displacement deviation threshold; when the maximum displacement deviation value is greater than the displacement deviation threshold, it is determined that the degree of asynchrony between the multiple cylinders has exceeded the safety boundary of not causing door jamming under the current differential pressure condition, and differentiated oil supply adjustment needs to be initiated; the control unit obtains the displacement data of each hydraulic cylinder, marks the hydraulic cylinder with the largest current extension displacement as the leading cylinder, marks the hydraulic cylinder with the smallest current extension displacement as the lagging cylinder, determines the proportional valve responsible for supplying hydraulic oil to the leading cylinder as the first proportional valve, and determines the proportional valve responsible for supplying hydraulic oil to the lagging cylinder as the second proportional valve.

[0020] Step 105: Calculate the difference between the maximum displacement deviation value and the displacement deviation threshold; based on the difference and the internal and external pressure difference value, obtain the flow compensation coefficient.

[0021] Specifically, the control unit calculates the difference between the maximum displacement deviation value and the displacement deviation threshold. The magnitude of this difference reflects the extent to which the asynchrony of the multiple cylinders exceeds the safety boundary; a larger difference indicates a greater need for compensation. Based on this difference and the internal and external pressure difference, the control unit derives a flow compensation coefficient. This flow compensation coefficient is used to determine the adjustment range for the oil supply to the leading and lagging cylinders; a larger difference or a larger internal and external pressure difference results in a larger flow compensation coefficient.

[0022] Step 106: Based on the flow compensation coefficient, generate a first oil supply command for controlling the first proportional valve, and / or generate a second oil supply command for controlling the second proportional valve; the first oil supply command is used to reduce the oil supply to the leading cylinder, and the second oil supply command is used to increase the oil supply to the lagging cylinder. Specifically, the control unit generates the oil supply command based on the flow compensation coefficient; for the first proportional valve, the control unit multiplies the current basic oil supply command by a proportional factor less than one determined by the flow compensation coefficient to generate the first oil supply command, so as to reduce the oil supply to the leading cylinder and slow down the extension speed of the leading cylinder; for the second proportional valve, the control unit multiplies the current basic oil supply command by a proportional factor greater than one determined by the flow compensation coefficient to generate the second oil supply command, so as to increase the oil supply to the lagging cylinder and speed up the extension speed of the lagging cylinder; through the above differential adjustment, the lagging cylinder gradually catches up with the leading cylinder, and the maximum displacement deviation value gradually decreases to within the displacement deviation threshold.

[0023] The above technical solution calculates the displacement deviation threshold based on the internal and external pressure difference. Under deep-water high-pressure differential conditions, the synchronization control precision is tightened to prevent door jamming and seal damage. Under shallow-water low-pressure differential conditions, the synchronization control precision is relaxed to increase the door opening speed. This solves the technical problem that fixed preset values ​​cannot adapt to the variable working conditions in deep-sea environments. By determining the advance cylinder and the lag cylinder and making targeted differentiated oil supply adjustments, unnecessary large-scale synchronization correction operations are reduced, improving the operating efficiency of the hydraulic system.

[0024] In other embodiments, the environmental water pressure data may not be collected directly by external pressure sensors, but rather calculated based on the current diving depth of the diving bell and the density of the seawater; the chamber water pressure data may not be collected by internal pressure sensors, but rather estimated based on the operating parameters of the diving bell's inflation and deflation system.

[0025] In some other embodiments, the maximum displacement deviation between multiple hydraulic cylinders may not be calculated by the difference between the maximum and minimum values, but rather by calculating the maximum deviation between the displacement data of each hydraulic cylinder and the average displacement value, or by calculating the maximum value among the absolute values ​​of the differences between the displacement data of any two hydraulic cylinders.

[0026] In some other embodiments, the flow compensation coefficient may not be obtained by direct calculation based on the difference and the internal and external pressure difference, but by looking up a pre-stored flow compensation coefficient mapping table, which records the compensation coefficient values ​​corresponding to different difference ranges and internal and external pressure difference ranges.

[0027] In some embodiments, calculating the displacement deviation threshold based on the internal and external pressure difference includes: The control unit acquires the hatch stiffness value and a preset initial tolerance value, which characterizes the initial allowable displacement deviation of the hatch under zero differential pressure. Specifically, the control unit reads the pre-stored hatch stiffness value; the hatch stiffness value reflects the hatch structure's resistance to deformation under stress, and this value can be pre-calibrated through finite element analysis or physical mechanical testing. Simultaneously, the control unit reads the preset initial tolerance value, which refers to the hydraulic cylinder displacement deviation that the hatch structure itself can accommodate when the diving bell is at the water surface, i.e., in a zero differential pressure environment.

[0028] Based on the hatch stiffness value and the internal and external pressure difference value, the stiffness-pressure difference ratio is obtained. Specifically, the control unit divides the hatch stiffness value by the internal and external pressure difference value to obtain the stiffness-pressure difference ratio. The stiffness-pressure difference ratio reflects the relationship between the hatch's own stiffness and the external pressure load. When the stiffness-pressure difference ratio is small, it indicates that the external pressure difference dominates relative to the hatch stiffness, and the hatch is prone to deformation under water pressure. At this time, the hatch's tolerance to cylinder asynchrony is reduced. When the stiffness-pressure difference ratio is large, it indicates that the hatch's own stiffness dominates relative to the external pressure difference, and the hatch's deformation under water pressure is smaller, with a relatively higher tolerance to cylinder asynchrony.

[0029] Step 203: Based on the initial tolerance value and the stiffness-pressure difference ratio, obtain the displacement deviation threshold. Specifically, the control unit divides the initial tolerance value by the stiffness-pressure difference ratio to obtain the displacement deviation threshold. When the stiffness-pressure difference ratio is small, i.e., the external pressure difference dominates, the displacement deviation threshold decreases accordingly, i.e., the synchronization control accuracy is tightened. When the stiffness-pressure difference ratio is large, i.e., the door stiffness dominates, the displacement deviation threshold increases accordingly, and the synchronization control accuracy can be appropriately relaxed.

[0030] In some embodiments, determining a leading cylinder and a lagging cylinder among the plurality of hydraulic cylinders, and determining a proportional valve supplying oil to the leading cylinder as a first proportional valve, and a proportional valve supplying oil to the lagging cylinder as a second proportional valve, includes: Obtain the telescopic displacement values ​​of the plurality of hydraulic cylinders in the corresponding motion direction.

[0031] The hydraulic cylinder with the largest telescopic displacement value is marked as the leading cylinder, and the proportional valve supplying oil to the leading cylinder is identified as the first proportional valve. Specifically, the control unit iterates through the telescopic displacement values ​​of all hydraulic cylinders, identifies the hydraulic cylinder with the largest telescopic displacement value, and marks it as the leading cylinder; at the same time, the control unit finds the proportional valve corresponding to the leading cylinder according to the one-to-one correspondence between proportional valves and hydraulic cylinders, and identifies it as the first proportional valve.

[0032] The hydraulic cylinder with the smallest telescopic displacement value is designated as the hysteresis cylinder, and the proportional valve supplying oil to the hysteresis cylinder is designated as the second proportional valve. Specifically, the hydraulic cylinder with the smallest telescopic displacement value is designated as the hysteresis cylinder, and the proportional valve corresponding to the hysteresis cylinder is designated as the second proportional valve.

[0033] By using the above technical solution, the advanced cylinder and the lagging cylinder are determined by the maximum and minimum values, which reduces the amount of calculation in the control unit in each control cycle and reduces the impact of control delay on the accuracy of synchronous control.

[0034] In other embodiments, the determination of the leading and lagging cylinders may not be limited to the two endpoints of maximum and minimum. Instead, all hydraulic cylinders are sorted from largest to smallest displacement value, and the M cylinders with the largest displacement are determined as the leading cylinder group and the K cylinders with the smallest displacement are determined as the lagging cylinder group, so as to realize the group control of multiple cylinders. M and K are preset positive integers and M plus K is less than the total number of hydraulic cylinders.

[0035] In some other embodiments, a minimum deviation dead zone interval can be set. When the maximum displacement deviation value falls within this dead zone interval, the determination operation of the advance and lag cylinders is not performed, thus avoiding frequent adjustments in scenarios with small deviations.

[0036] In some embodiments, generating a first oil supply command for controlling the first proportional valve, and / or generating a second oil supply command for controlling the second proportional valve includes: The current basic command of the first proportional valve and / or the current basic command of the second proportional valve are obtained. Specifically, the control unit reads the oil supply command value of the currently executing first proportional valve and / or second proportional valve from the command buffer as the current basic command; the current basic command reflects the valve core opening control signal value of each proportional valve before the execution of this synchronous correction adjustment, and this value is calculated from the command of the previous control cycle and continues to be effective in the current control cycle.

[0037] Based on the current basic command of the first proportional valve and the flow compensation coefficient, the first oil supply command is obtained. Specifically, the control unit uses the current basic command of the first proportional valve as a reference, converts the flow compensation coefficient into an attenuation factor less than one, and multiplies the current basic command by the attenuation factor to obtain the first oil supply command. If the value of the first oil supply command is lower than the current basic command, the corresponding proportional valve spool opening decreases, thereby reducing the oil supply flow to the advance cylinder.

[0038] Based on the current basic command of the second proportional valve and the flow compensation coefficient, the second oil supply command is obtained. Specifically, the control unit uses the current basic command of the second proportional valve as a reference, converts the flow compensation coefficient into an enhancement factor greater than one, and multiplies the current basic command by the enhancement factor to obtain the second oil supply command. The value of the second oil supply command is higher than that of the current basic command, and the corresponding proportional valve spool opening increases, thereby increasing the oil supply flow to the lag cylinder.

[0039] By using the above technical solution, adjustments are made based on the current basic commands of each proportional valve, which reduces the disturbance to the control signal of the entire hydraulic circuit, making the transition of control commands smoother and more continuous, and reducing the pressure shock and oscillation of hydraulic pipelines caused by large jumps in control commands.

[0040] In some other embodiments, the generation of the first oil supply command and the second oil supply command is not performed independently. Instead, while keeping the total oil supply constant, the oil supply reduced by the leading oil cylinder is proportionally allocated to the lagging oil cylinder to achieve a redistribution of the oil supply.

[0041] In some embodiments, the method further includes: Obtain a preset off-center load coefficient mapping table, which records the correspondence between different stroke ranges and the off-center load weight coefficients of each hydraulic cylinder.

[0042] Specifically, the control unit reads a pre-calibrated off-center load factor mapping table. This mapping table is established during the factory calibration or maintenance calibration phase by testing and calculating the center of gravity position of the diving bell hatch at different opening angles. During the opening process, the center of gravity of the hatch continuously shifts, and the equivalent load borne by different hydraulic cylinders varies with the hatch opening angle. The off-center load factor mapping table uses the hatch stroke as an index to record the off-center load weight coefficient of each hydraulic cylinder within different stroke ranges. This coefficient reflects the proportion of the equivalent load caused by gravity off-center loading for each hydraulic cylinder at the current stroke.

[0043] The average displacement data of the multiple hydraulic cylinders is calculated to obtain the current stroke. Specifically, the control unit sums the current displacement data of each hydraulic cylinder and divides it by the total number of hydraulic cylinders to obtain the average displacement value of each hydraulic cylinder, which is used as the current stroke of the hatch. The current stroke refers to the displacement of the hatch from the fully closed position to the current position, which is used to characterize the current opening degree of the hatch.

[0044] Based on the current stroke, the control unit looks up the corresponding off-center load coefficient mapping table to extract the first off-center load weight coefficient corresponding to the leading cylinder and the second off-center load weight coefficient corresponding to the lagging cylinder. Specifically, the control unit uses the current stroke as an index to search for the corresponding stroke interval in the off-center load coefficient mapping table, reads the off-center load weight coefficient corresponding to the leading cylinder in that stroke interval as the first off-center load weight coefficient, and reads the off-center load weight coefficient corresponding to the lagging cylinder as the second off-center load weight coefficient. If the current stroke falls exactly between two stroke intervals, the control unit can use linear interpolation of adjacent interval data to determine the corresponding off-center load weight coefficient.

[0045] Based on the product of the flow compensation coefficient and the first off-center load weighting coefficient, the current basic command of the first proportional valve is corrected to obtain the final first oil supply command. Specifically, the control unit multiplies the flow compensation coefficient by the first off-center load weighting coefficient to obtain the corrected flow compensation coefficient; the control unit uses the current basic command of the first proportional valve as a reference, converts the corrected flow compensation coefficient into a decay factor, and multiplies it by the current basic command to obtain the final first oil supply command. When the first off-center load weighting coefficient is small, it indicates that the gravity off-center load on the advance cylinder is small in the current stroke, and its fast movement is mainly due to the uneven hydraulic oil supply. At this time, it is necessary to increase the deceleration force. Therefore, the product of the first off-center load weighting coefficient and the flow compensation coefficient further increases the deceleration amplitude of the advance cylinder.

[0046] Based on the product of the flow compensation coefficient and the second off-center load weight coefficient, the current basic command of the second proportional valve is corrected to obtain the final second oil supply command.

[0047] Specifically, the control unit multiplies the flow compensation coefficient by the second off-center load weighting coefficient and converts it into an enhancement factor. This factor is then multiplied by the current base command of the second proportional valve to obtain the final second oil supply command. When the second off-center load weighting coefficient is large, it indicates that the lagging cylinder experiences a large gravitational off-center load during its current stroke. A significant portion of its slow movement is due to structural off-center load rather than insufficient hydraulic oil supply. In this case, increased acceleration is needed to overcome the off-center load resistance. Therefore, the product of the second off-center load weighting coefficient and the flow compensation coefficient results in a corresponding increase in the acceleration amplitude of the lagging cylinder.

[0048] Through the above technical solution, the gravity load on different hydraulic cylinders varies at different stages of the door stroke. By looking up the corresponding load weight coefficient, the flow compensation is corrected, thereby realizing oil supply control under asymmetrical load and reducing synchronous control failure and cylinder reciprocating oscillation caused by load imbalance.

[0049] In other embodiments, the off-center load coefficient mapping table may not use a segmented interval lookup table, but instead be calculated as a continuous function of the pre-fitted off-center load weight coefficient with respect to the travel.

[0050] In some embodiments, after calculating the displacement deviation threshold, the method further includes: The upper and lower limits of the allowable deviation are obtained. Specifically, the control unit's memory has two preset parameters: the upper and lower limits of the allowable deviation. The upper limit of the allowable deviation represents the maximum allowable value of the displacement deviation threshold. Exceeding this limit means that the synchronization control precision is too lenient, which may cause significant deviation of the hatch movement and affect the sealing performance. The lower limit of the allowable deviation represents the minimum allowable value of the displacement deviation threshold. Below this limit means that the synchronization control precision requirement is too strict, which may cause frequent adjustments to the hydraulic system, resulting in unnecessary energy consumption and shocks.

[0051] When the displacement deviation threshold is greater than the upper limit of the allowable deviation, the displacement deviation threshold is set to the upper limit of the allowable deviation. Specifically, the control unit compares the displacement deviation threshold with the upper limit of the allowable deviation; if the displacement deviation threshold exceeds the upper limit of the allowable deviation, the control unit forcibly sets the displacement deviation threshold to the value of the upper limit of the allowable deviation to prevent the displacement deviation threshold from being too large under extremely low differential pressure conditions, which would lead to low synchronization control accuracy and cause significant skewness in the hatch operation.

[0052] When the displacement deviation threshold is less than the lower limit of the allowable deviation, the displacement deviation threshold is set to the lower limit of the allowable deviation. Specifically, the control unit compares the displacement deviation threshold with the lower limit of the allowable deviation; if the displacement deviation threshold is lower than the lower limit of the allowable deviation, the control unit sets the displacement deviation threshold to the value of the lower limit of the allowable deviation to prevent the hydraulic system from making unnecessary rapid adjustments within a small deviation range due to an excessively small displacement deviation threshold under extremely high pressure differential conditions.

[0053] The above technical solution limits the upper and lower limits of the displacement deviation threshold calculated from the door stiffness value and the internal and external pressure difference value, so as to avoid the calculated value from exceeding the reasonable range under extreme working conditions.

[0054] In some embodiments, after presetting the upper limit and lower limit of the allowable deviation, and before setting the displacement deviation threshold, the method further includes: Obtain the maximum working pressure difference value corresponding to the maximum designed working depth of the diving bell. Specifically, the control unit reads the maximum working pressure difference value corresponding to the maximum designed working depth of the diving bell from the memory; the maximum working pressure difference value is the limit of the internal and external pressure difference that the diving bell needs to withstand at the deepest working depth specified in the diving bell's technical specifications.

[0055] Based on the current internal and external pressure difference and the maximum working pressure difference, the depth load ratio is obtained. Specifically, the control unit divides the current internal and external pressure difference by the maximum working pressure difference to obtain the depth load ratio. The depth load ratio reflects the current relative position of the diving bell within its designed working depth range. A depth load ratio close to one indicates that the current depth is at or near the maximum working depth of the diving bell, while a depth load ratio relatively small indicates that the current depth is in shallower water.

[0056] When the depth-load ratio is greater than a preset high-load depth threshold, the upper limit of the allowable deviation is reduced based on the upper limit of the allowable deviation and a preset first reduction ratio coefficient, and the lower limit of the allowable deviation is reduced based on the lower limit of the allowable deviation and a preset second reduction ratio coefficient. Specifically, when the depth-load ratio is greater than the preset high-load depth threshold, i.e., under high-depth extreme conditions, to ensure the safe operation of the hatch, the control unit multiplies the upper limit of the allowable deviation by a first reduction ratio coefficient less than one to obtain a reduced upper limit of the allowable deviation, and multiplies the lower limit of the allowable deviation by a second reduction ratio coefficient less than one to obtain a reduced lower limit of the allowable deviation. This forcibly tightens the limiting range, preventing small displacement deviations from being ignored due to an excessively wide limiting range of the displacement deviation threshold under extreme conditions, thus increasing protection against irreversible structural damage under high pressure.

[0057] When the depth-load ratio is less than a preset depth threshold, the upper limit of the allowable deviation is amplified based on the upper limit of the allowable deviation and a preset first amplification ratio coefficient, and the lower limit of the allowable deviation is amplified based on the lower limit of the allowable deviation and a preset second amplification ratio coefficient. Specifically, when the depth-load ratio is less than the preset depth threshold, i.e., in a shallow water low-load condition, the control unit multiplies the upper limit of the allowable deviation by a first amplification ratio coefficient greater than one to obtain an amplified upper limit of the allowable deviation, and multiplies the lower limit of the allowable deviation by a second amplification ratio coefficient greater than one to obtain an amplified lower limit of the allowable deviation, thereby widening the limiting range, allowing operation with a larger displacement deviation under shallow water safety conditions, reducing unnecessary adjustment frequency of the proportional valve, and increasing the opening speed.

[0058] The above technical solution allows for scaling and adjustment of the upper and lower limits of allowable deviation based on depth load comparison, enabling the limiting range to change according to the current operating depth of the diving bell. In extreme deep-water conditions, the tolerance value is tightened to ensure structural safety, while in shallow-water low-pressure conditions, the tolerance value is released to improve operating efficiency.

[0059] In other embodiments, the first reduction ratio and the second reduction ratio may not be fixed, but may be set in segments according to the depth load ratio. The closer the depth load ratio is to one, the smaller the reduction ratio.

[0060] In some embodiments, after calculating the flow compensation coefficient, the method further includes: The historical flow compensation coefficient from the previous calculation cycle is retrieved. Specifically, the control unit reads the flow compensation coefficient calculated in the previous control cycle from the internal buffer as the historical flow compensation coefficient. At the end of each control cycle, the control unit stores the flow compensation coefficient calculated in the current cycle into the buffer for use in the next cycle.

[0061] Calculate the absolute value of the compensation deviation between the current flow compensation coefficient and the historical flow compensation coefficient. Specifically, the control unit subtracts the historical flow compensation coefficient from the flow compensation coefficient calculated in the current cycle and takes the absolute value to obtain the absolute value of the compensation deviation. The absolute value of the compensation deviation reflects the magnitude of the change in the flow compensation coefficient between adjacent control cycles; the larger the value, the more drastic the jump in compensation intensity.

[0062] When the absolute value of the compensation deviation is greater than a preset hydraulic allowable step size, the control unit determines the relationship between the current flow compensation coefficient and the historical flow compensation coefficient. Specifically, the control unit compares the absolute value of the compensation deviation with the preset hydraulic allowable step size; the hydraulic allowable step size is a safety limit set based on the maximum allowable pressure change rate of the hydraulic pipeline and the response characteristics of the proportional valve. Flow compensation command jumps exceeding this step size may cause water hammer or pressure shocks in the hydraulic pipeline. When the absolute value of the compensation deviation is greater than the hydraulic allowable step size, the control unit further determines the relationship between the current flow compensation coefficient and the historical flow compensation coefficient to determine whether an escalation step size limit or a de-escalation step size limit is required.

[0063] When the current flow compensation coefficient is greater than the historical flow compensation coefficient, the current flow compensation coefficient is reset to the sum of the historical flow compensation coefficient and the hydraulic allowable step size. Specifically, when the control unit determines that the current flow compensation coefficient is greater than the historical flow compensation coefficient, it indicates that the flow compensation coefficient has experienced a significant upward jump in this cycle. To prevent excessive flow compensation commands from exacerbating hydraulic pipeline shocks, the control unit resets the flow compensation coefficient to the sum of the historical flow compensation coefficient and the hydraulic allowable step size, achieving a gradual increase rather than a one-time large jump.

[0064] When the current flow compensation coefficient is less than the historical flow compensation coefficient, the current flow compensation coefficient is reset to the difference between the historical flow compensation coefficient and the hydraulic allowable step size. Specifically, when the control unit determines that the current flow compensation coefficient is less than the historical flow compensation coefficient, it indicates that the flow compensation coefficient has experienced a significant downward jump in this cycle; the control unit resets the flow compensation coefficient to the difference between the historical flow compensation coefficient and the hydraulic allowable step size, achieving a gradual decrease in steps.

[0065] In some embodiments, after resetting the traffic compensation coefficient, the method further includes: The real-time movement speed data of the multiple hydraulic cylinders is acquired. Specifically, the control unit calculates the real-time movement speed data of each hydraulic cylinder by dividing the displacement change of each displacement sensor at adjacent sampling times by the sampling time interval; alternatively, the control unit can directly acquire the real-time movement speed data through the speed sensors installed on each hydraulic cylinder.

[0066] Based on the real-time motion speed data and the preset rated speed, a speed ratio is obtained. Specifically, the control unit calculates the average of the real-time motion speed data of each hydraulic cylinder as the current motion speed, and divides the current motion speed by the preset rated speed to obtain the speed ratio; the preset rated speed is the designed motion speed of the hydraulic cylinder under rated oil supply conditions. A speed ratio greater than one indicates that the current motion speed is higher than the rated speed, i.e., it is in a fast motion state; a speed ratio less than one indicates that the current motion speed is lower than the rated speed, i.e., it is in a slow motion state.

[0067] Based on the flow compensation coefficient and the speed ratio, a corrected flow compensation coefficient is obtained. The larger the value corresponding to the real-time motion speed data, the larger the speed ratio, and the greater the corrected flow compensation coefficient. Specifically, the control unit multiplies the flow compensation coefficient after amplitude limiting by the speed ratio to obtain the corrected flow compensation coefficient. When the hatch operates at a higher speed, the displacement deviation between different cylinders develops faster, requiring a stronger flow compensation force to quickly correct the deviation. When the hatch operates at a lower speed, the deviation develops more slowly, and the flow compensation force can be reduced accordingly to avoid oscillations caused by overcompensation.

[0068] The above technical solution establishes a matching relationship between the compensation intensity and the actual movement speed of the hydraulic cylinder. It provides sufficient flow compensation to achieve rapid synchronous correction when the hatch is running at high speed, and reduces the compensation intensity to avoid oscillation caused by overcompensation when running at low speed. This solves the problem of insufficient adaptability of fixed compensation intensity under high and low speed conditions, which leads to overshoot or correction lag.

[0069] In some embodiments, after receiving the first fuel supply command and / or the second fuel supply command, the method further includes: Preset target position data is acquired, which represents the endpoint of the hatch's travel when it is fully open or fully closed. Specifically, the control unit reads the preset target position data from the memory; the target position data is written to the memory during the system calibration phase, recording the absolute displacement value corresponding to the endpoint of the hatch's travel from the fully closed position to the fully open position. Depending on the current direction of hatch movement, the target position data corresponds to either the endpoint of the hatch's travel when it is fully open or the endpoint of the hatch's travel when it is fully closed.

[0070] The remaining travel distance between the displacement data and the target position data is calculated. Specifically, the control unit subtracts the average value of the current displacement data of each hydraulic cylinder from the target position data to obtain the remaining travel distance. The remaining travel distance reflects how close the current position of the hatch is to the end of the travel; the smaller the remaining travel distance, the closer the hatch is to the end of the mechanical travel.

[0071] When the remaining travel distance is less than a preset buffer distance, an attenuation ratio is generated based on the remaining travel distance; the attenuation ratio is less than one. Specifically, when the remaining travel distance is less than the preset buffer distance, it indicates that the hatch is about to reach the end of its mechanical travel. The control unit generates an attenuation ratio based on the remaining travel distance; the value of this attenuation ratio is greater than zero and less than one, and the attenuation ratio changes in the same direction as the remaining travel distance, that is, the smaller the remaining travel distance, the smaller the attenuation ratio. For example, the attenuation ratio can be taken as the ratio of the remaining travel distance to the buffer distance.

[0072] Based on the attenuation ratio, the calculated first and / or second oil supply commands are adjusted. Specifically, the first and / or second oil supply commands obtained by the control unit are multiplied by the attenuation ratio to obtain the final oil supply command with reduced amplitude. As the hatch gradually approaches the end of its stroke, the attenuation ratio continues to decrease, and the amplitude of the oil supply command decreases synchronously, so that the movement speed of the hydraulic cylinder smoothly transitions to a stop at the end of its stroke, avoiding a violent impact on the hatch at the mechanical end point due to high-speed inertial impact.

[0073] The above technical solution avoids severe impact damage between the hatch and the frame caused by high-speed inertia, ensuring the smoothness of the hatch opening and closing process throughout.

[0074] All of the above-mentioned optional technical solutions can be combined in any way to form the optional embodiments of this application, and will not be described in detail here.

[0075] Those skilled in the art will recognize that the units and algorithm steps of the various examples described in conjunction with the embodiments disclosed herein can be implemented in electronic hardware, or a combination of computer software and electronic hardware. Whether these functions are implemented in hardware or software depends on the specific application and design constraints of the technical solution. Those skilled in the art can use different methods to implement the described functions for each specific application, but such implementation should not be considered beyond the scope of this application.

[0076] Those skilled in the art will understand that, for the sake of convenience and brevity, the specific working processes of the systems, devices, and units described above can be referred to the corresponding processes in the foregoing method embodiments, and will not be repeated here.

[0077] In the several embodiments provided in this application, it should be understood that the disclosed systems, apparatuses, and methods can be implemented in other ways. For example, the apparatus embodiments described above are merely illustrative; for instance, the division of units is only a logical functional division, and in actual implementation, there may be other division methods. For example, multiple units or components may be combined or integrated into another system, or some features may be ignored or not executed. Furthermore, the coupling or direct coupling or communication connection shown or discussed may be through some interfaces; the indirect coupling or communication connection between apparatuses or units may be electrical, mechanical, or other forms.

[0078] The units described as separate components may or may not be physically separate. The components shown as units may or may not be physical units; that is, they may be located in one place or distributed across multiple network units. Some or all of the units can be selected to achieve the purpose of this embodiment according to actual needs.

[0079] In addition, the functional units in the various embodiments of this application can be integrated into one processing unit, or each unit can exist physically separately, or two or more units can be integrated into one unit.

[0080] If the aforementioned functions are implemented as software functional units and sold or used as independent products, they can be stored in a computer-readable storage medium. Based on this understanding, the technical solution of this application, in essence, or the part that contributes to the prior art, or a portion of the technical solution, can be embodied in the form of a software product. This computer software product is stored in a storage medium and includes several instructions to cause a computer device (which may be a personal computer, server, or network device, etc.) to execute all or part of the steps of the methods described in the various embodiments of this application. The aforementioned storage medium includes various media capable of storing program verification codes, such as USB flash drives, portable hard drives, read-only memory (ROM), random access memory (RAM), magnetic disks, or optical disks.

[0081] In the description of this application, the terms "first" and "second" are used for descriptive purposes only and should not be construed as indicating or implying relative importance or implicitly specifying the number of technical features indicated. Therefore, a feature defined as "first" or "second" may explicitly or implicitly include one or more features. In the description of this application, "multiple" means two or more, unless otherwise explicitly specified.

[0082] In the above embodiments, the descriptions of each embodiment have different focuses. For parts not described in detail in a certain embodiment, please refer to the relevant descriptions in other embodiments.

[0083] The embodiments, implementation methods, and related technical features of this application can be combined and substituted for each other without conflict.

[0084] The above are merely preferred embodiments of this application and are not intended to limit this application in any way. Any simple modifications, equivalent changes, and alterations made to the above embodiments based on the technical essence of this application without departing from the scope of the technical solution of this application shall still fall within the scope of the technical solution of this application.

Claims

1. A method for synchronous control of multiple hydraulic cylinders in a diving bell hatch, characterized in that, include: Acquire environmental water pressure data outside the diving bell and water pressure data inside the chambers; Calculate the difference between the ambient water pressure data and the cabin water pressure data to obtain the internal and external pressure difference value; Obtain displacement data of multiple hydraulic cylinders that drive the opening of the hatch; Calculate the maximum displacement deviation value between the plurality of hydraulic cylinders based on the displacement data; The displacement deviation threshold is calculated based on the internal and external pressure difference value. The displacement deviation threshold represents the maximum allowable displacement difference that does not cause door jamming under the current pressure difference condition. The internal and external pressure difference value is inversely proportional to the displacement deviation threshold. When the maximum displacement deviation value is greater than the displacement deviation threshold, the displacement data of the plurality of hydraulic cylinders are compared to determine the leading cylinder and the lagging cylinder among the plurality of hydraulic cylinders; the proportional valve that supplies oil to the leading cylinder is determined as the first proportional valve, and the proportional valve that supplies oil to the lagging cylinder is determined as the second proportional valve. Calculate the difference between the maximum displacement deviation value and the displacement deviation threshold; Based on the difference and the internal and external pressure difference, the flow compensation coefficient is obtained; Based on the flow compensation coefficient, a first oil supply command is generated to control the first proportional valve, and / or a second oil supply command is generated to control the second proportional valve; the first oil supply command is used to reduce the oil supply to the leading cylinder, and the second oil supply command is used to increase the oil supply to the lagging cylinder.

2. The method according to claim 1, characterized in that, The displacement deviation threshold is calculated based on the aforementioned internal and external pressure difference, including: Obtain the hatch stiffness value and the preset initial tolerance value; the initial tolerance value represents the initial allowable displacement deviation of the hatch under zero differential pressure environment; Based on the door stiffness value and the internal and external pressure difference value, the stiffness-pressure difference ratio is obtained; The displacement deviation threshold is obtained based on the initial tolerance value and the stiffness-pressure difference ratio.

3. The method according to claim 2, characterized in that, Identifying the leading cylinder and the lagging cylinder among the plurality of hydraulic cylinders, and designating the proportional valve supplying oil to the leading cylinder as the first proportional valve, and the proportional valve supplying oil to the lagging cylinder as the second proportional valve, includes: Obtain the extension and retraction displacement values ​​of the plurality of hydraulic cylinders in the corresponding motion direction; The hydraulic cylinder with the largest telescopic displacement value is marked as the advance cylinder, and the proportional valve that supplies oil to the advance cylinder is determined as the first proportional valve. The hydraulic cylinder with the smallest extension displacement value is designated as the hysteresis cylinder, and the proportional valve that supplies oil to the hysteresis cylinder is designated as the second proportional valve.

4. The method according to claim 3, characterized in that, Generating a first oil supply command for controlling the first proportional valve, and / or generating a second oil supply command for controlling the second proportional valve, including: Obtain the current basic command of the first proportional valve and / or the current basic command of the second proportional valve; Based on the current basic command of the first proportional valve and the flow compensation coefficient, the first oil supply command is obtained; And / or, based on the current base command of the second proportional valve and the flow compensation coefficient, the second oil supply command is obtained.

5. The method according to claim 4, characterized in that, The method further includes: Obtain a preset off-center load coefficient mapping table; the off-center load coefficient mapping table records the correspondence between different stroke ranges and the off-center load weight coefficients of each hydraulic cylinder; Calculate the average value of the displacement data of the multiple hydraulic cylinders to obtain the current stroke; According to the current stroke, the table is looked up in the off-center load coefficient mapping table to extract the first off-center load weight coefficient corresponding to the leading cylinder and the second off-center load weight coefficient corresponding to the lagging cylinder; The current basic command of the first proportional valve is corrected based on the product of the flow compensation coefficient and the first off-center load weight coefficient to obtain the final first oil supply command. The current basic command of the second proportional valve is corrected based on the product of the flow compensation coefficient and the second off-center load weight coefficient to obtain the final second oil supply command.

6. The method according to claim 5, characterized in that, After calculating the displacement deviation threshold, the method further includes: Obtain the upper and lower limits of the allowable deviation; When the displacement deviation threshold is greater than the upper limit of the allowable deviation, the displacement deviation threshold is set to the upper limit of the allowable deviation; When the displacement deviation threshold is less than the allowable deviation lower limit, the displacement deviation threshold is set as the allowable deviation lower limit.

7. The method according to claim 6, characterized in that, The method further includes: Obtain the maximum working pressure difference value corresponding to the maximum working depth of the diving bell design; Based on the current internal and external pressure difference value and the maximum working pressure difference value, the depth load ratio is obtained; When the depth-to-load ratio is greater than a preset high-load depth threshold, the upper limit of the allowable deviation is reduced based on the upper limit of the allowable deviation and a preset first reduction ratio coefficient; the lower limit of the allowable deviation is reduced based on the lower limit of the allowable deviation and a preset second reduction ratio coefficient. When the depth-load ratio is less than a preset depth threshold, the upper limit of the allowable deviation is amplified based on the upper limit of the allowable deviation and a preset first amplification ratio coefficient; the lower limit of the allowable deviation is amplified based on the lower limit of the allowable deviation and a preset second amplification ratio coefficient.

8. The method according to claim 7, characterized in that, The method further includes: Obtain the real-time movement speed data of the plurality of hydraulic cylinders; Based on the real-time motion speed data and the preset rated speed, the speed ratio is obtained; Based on the flow compensation coefficient and the speed ratio, a corrected flow compensation coefficient is obtained; wherein, the larger the value corresponding to the real-time motion speed data, the larger the speed ratio, and the corrected flow compensation coefficient increases accordingly.

9. The method according to claim 8, characterized in that, The method further includes: Acquire target position data; the target position data represents the end point of the journey when the hatch is fully open or fully closed; Calculate the remaining travel distance between the displacement data and the target position data; When the remaining travel distance is less than a preset buffer distance, an attenuation ratio is generated based on the remaining travel distance; the attenuation ratio is less than one. Based on the attenuation ratio, the calculated first fuel supply command and / or second fuel supply command are adjusted.

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