A rotary pump integrated multi-functional oil-gas mixed transportation system

Abstract

This invention relates to the field of oil and gas mixed transportation equipment technology, and discloses a rotor pump integrated multifunctional oil and gas mixed transportation system, including a dual-chamber body, pipelines, sensors, actuators, and a main control module. The dual-chamber body is divided into a first chamber and a second chamber. A bypass pipeline connects to the fluid inlet and outlet of the first chamber and is connected in series with a throttling and pressure-reducing valve. The outlet of the throttling and pressure-reducing valve is connected to a heat exchange jacket covering the outside of the second chamber. The main control module calculates the airflow velocity quotient based on the pressure parameters of the first chamber. When the quotient is exceeded, a servo regulating valve is opened to discharge gas into the bypass pipeline. After being throttled, the gas expands and cools down before entering the heat exchange jacket to cool the oil in the second chamber. Simultaneously, a dynamic heat load index is generated by combining the mechanical seal temperature, temperature rise rate, and flow velocity quotient, and the speed of the oil circulation pump is adaptively adjusted to lubricate and cool the mechanical seal. This invention directly transforms the exhaust unloading condition into an internal cooling source, achieving advanced temperature intervention on the mechanical seal and preventing thermal damage caused by dry friction.

Description

Technical Field

[0001] This invention relates to the field of oil and gas mixed transportation equipment technology, specifically to a rotor pump integrated multi-functional oil and gas mixed transportation system. Background Technology

[0002] In oilfield extraction and gathering, the gas-liquid ratio of the produced fluid often fluctuates significantly. Multiphase flow rotor pumps are commonly used for mixed oil and gas transportation. When transporting gaseous media, rotor pumps primarily rely on the liquid remaining in the pump chamber to maintain the liquid seal between the rotors and lubricate the mechanical seal components. In actual operation, multiphase fluids often exhibit slug flow characteristics. When a large amount of gas rushes into the pump chamber instantaneously, the kinetic energy of the gas flow inside the fluid increases, disrupting the liquid surface tension and causing the lubricating liquid remaining in the chamber to be carried away by the gas flow. This liquid loss directly alters the working environment of the mechanical seal components, inducing dry friction and generating frictional heat.

[0003] To address the overheating issue of mechanical seals, existing multiphase flow pumps often employ an external cooling water system in conjunction with a single closed-loop temperature feedback control. This control method relies on changes in absolute temperature, but the inherent time lag in heat conduction of solid media means that by the time the control system detects a temperature rise and issues a cooling command, the mechanical seal has already undergone a period of thermal shock, increasing the risk of seal wear or failure. Furthermore, existing systems fail to establish a direct data correlation between the aerodynamic changes within the pump chamber and the thermal state of the mechanical seal, making it impossible to proactively intervene and compensate for cooling in the early stages of flow field instability and dry friction. Simultaneously, the reliance on external cooling sources increases the overall energy consumption and piping complexity of the system, making it difficult to apply in remote gathering and transportation nodes lacking auxiliary water sources. Summary of the Invention

[0004] To address the shortcomings of existing technologies, this invention provides an integrated multi-functional oil-gas mixed transport system for rotor pumps. This system solves the problems of dry friction caused by lubrication loss in mechanical seals of existing mixed transport pumps under high gas content conditions, and the thermal damage to equipment caused by the lag in traditional single temperature feedback control.

[0005] To achieve the above objectives, the present invention provides the following technical solution:

[0006] This invention provides an integrated multi-functional oil-gas mixed transportation system for a rotor pump, comprising a dual-chamber body, a pipeline module, a sensing module, an execution module, and a main control module;

[0007] The dual-cavity body is internally divided into a physically isolated first cavity and a second cavity; the first cavity contains an internal rotor, and the second cavity contains a mechanical seal assembly connected to the internal rotor;

[0008] The piping module includes a bypass line, a throttling and pressure-reducing valve, and a heat exchange jacket; the bypass line is connected between the fluid inlet and outlet of the first chamber; the throttling and pressure-reducing valve is connected in series in the bypass line; the heat exchange jacket covers the outside of the second chamber, the bypass line passes through the throttling and pressure-reducing valve and is connected to the fluid inlet end of the heat exchange jacket, and the fluid outlet end of the heat exchange jacket is connected to the fluid inlet end of the first chamber.

[0009] The sensing module includes a static pressure sensor, a dynamic pressure transmitter, an inlet temperature sensor, and a thermocouple assembly; the static pressure sensor, the dynamic pressure transmitter, and the inlet temperature sensor are assembled at the fluid inlet and outlet of the first chamber, and the thermocouple assembly is installed inside the mechanical seal assembly in the second chamber.

[0010] The execution module includes a servo regulating valve, an oil circulation pump, and a frequency converter; the servo regulating valve is connected in series in the bypass line and is located between the fluid outlet end of the first chamber and the throttle pressure reducing valve; the oil circulation pump is arranged in the oil circulation loop of the second chamber; and the frequency converter is connected to the oil circulation pump.

[0011] The main control module is connected to both the sensing module and the execution module.

[0012] Furthermore, the main control module is configured to perform the following working steps:

[0013] The sensor module reads the pressure and differential pressure values ​​inside the first chamber, and also reads the absolute temperature value of the mechanical seal assembly.

[0014] The apparent gas velocity and critical gas-liquid suspension velocity in the first chamber are calculated based on the pressure and differential pressure values. When the quotient of the apparent gas velocity and the critical gas-liquid suspension velocity reaches the trigger threshold, the servo regulating valve is opened to guide the gas in the first chamber to the bypass pipeline.

[0015] The gas in the bypass pipeline is guided to undergo adiabatic expansion and cooling through the throttling and pressure reducing valve. The cooled gas is then introduced into the heat exchange jacket to exchange heat with the oil medium inside the second chamber. After absorbing heat, the gas flows out of the heat exchange jacket and merges into the fluid inlet of the first chamber.

[0016] The derivative of temperature with respect to time is calculated based on the absolute temperature value, and the quotient, absolute temperature value, and derivative of temperature with respect to time are combined to generate a dynamic heat load index.

[0017] The frequency adjustment signal generated based on the dynamic heat load index is sent to the variable frequency drive to adjust the speed of the oil circulation pump, so as to circulate the cooled oil medium in the heat exchange jacket to the surface of the mechanical seal assembly, thereby realizing pneumatic unloading and scheduling and circulation of the cold source.

[0018] Furthermore, the main control module calculates the transient gas phase density in the first cavity based on the ideal gas state theory. The value of the transient gas phase density is equal to the transient static pressure at the airflow inlet obtained by the static pressure sensor multiplied by the molar mass of the transported gas, and then divided by the product of the universal gas constant and the absolute temperature of the intake airflow obtained by the inlet temperature sensor.

[0019] Furthermore, the process of the main control module calculating the apparent gas velocity is as follows: the value of the apparent gas velocity is equal to twice the pressure difference between the inlet and outlet of the first cavity fluid obtained by the dynamic pressure transmitter, divided by the product of the comprehensive resistance coefficient of the flow channel inside the first cavity and the transient gas density, and then the square root is taken.

[0020] Furthermore, the process of the main control module calculating the critical suspension velocity of gas and liquid is as follows: calculate the product of the surface tension at the gas-liquid two-phase contact interface and the difference between the liquid phase density and the transient gas phase density of the working liquid, divide it by the square of the transient gas phase density, raise the quotient to the power of 0.25, and then multiply it by the empirical constant characterizing the flow resistance and geometry inside the first cavity.

[0021] Furthermore, the quotient is the dimensionless dynamic anti-entrapment safety factor obtained by the main control module dividing the apparent gas velocity by the critical gas-liquid suspension velocity. When the dynamic anti-entrapment safety factor reaches or exceeds the trigger threshold, the main control module outputs a continuous analog opening control signal to the servo regulating valve. The opening of the servo regulating valve increases proportionally with the difference between the safety factor and the trigger threshold. This reduces the gas flow rate across the main liquid surface and prevents the loss of lubricating medium when the gas kinetic energy approaches the state of breaking the liquid seal.

[0022] Furthermore, the main control module incorporates a quantitative thermodynamic evaluation model to predict the available cold source quality at the outlet of the throttling and pressure reducing valve. The absolute temperature of the low-temperature gas after throttling and expansion is equal to the absolute temperature of the high-pressure gas before throttling, minus the product of the Joule-Thomson coefficient of the fluid medium and the static pressure difference between the inlet and outlet of the throttling and pressure reducing valve. This utilizes the adiabatic isenthalpic throttling effect to convert the discharged high-pressure gas flow into a cooling medium that reduces the heat generation state.

[0023] Furthermore, the main control module uses a backward differential algorithm to calculate the derivative of temperature with respect to time. This derivative value is equal to the difference between the absolute temperature value of the current sampling period and the absolute temperature value of the previous sampling period, divided by the interval time between two adjacent temperature samplings set by the system.

[0024] Furthermore, the main control module calculates the dynamic heat load index as follows: the dynamic heat load index is equal to the sum of the first, second, and third items; the first item is the ratio of the absolute temperature value of the mechanical seal assembly in the current cycle to the reference temperature under normal operating conditions, multiplied by the set steady-state weighting coefficient; the second item is the ratio of the temperature derivative with respect to time to the reference value of the maximum allowable temperature rise rate, multiplied by the set dynamic weighting coefficient; the third item is the quotient of the apparent gas flow rate and the critical gas-liquid suspension flow rate, multiplied by the set feedforward weighting coefficient, to provide an adjustment basis by integrating the steady-state temperature value, transient temperature rise rate, and aerodynamic feedforward.

[0025] Furthermore, the steady-state weight coefficient, dynamic weight coefficient, and feedforward weight coefficient are all dimensionless constants, and the sum of the three weight coefficients in the main control module is constrained to be one.

[0026] Furthermore, the main control module internally constructs an exponential adaptive adjustment function to calculate the target operating frequency corresponding to the frequency adjustment signal; when the dynamic heat load index is less than or equal to the set safe heat load threshold, the target operating frequency is equal to the minimum operating frequency required for the oil circulation pump to maintain basic lubrication; when the dynamic heat load index is greater than the safe heat load threshold, the target operating frequency is equal to the minimum operating frequency, plus the product of the nonlinear response exponent of the difference between the dynamic heat load index and the safe heat load threshold and the gain coefficient of the frequency control.

[0027] Furthermore, when the dynamic heat load index shows a downward trend and gradually decreases, the main control module synchronously lowers the target operating frequency through an exponential adaptive adjustment function to smoothly reduce the speed of the oil circulation pump, thereby avoiding excessive cooling of the working medium over a long period of time and an increase in non-power consumption.

[0028] Furthermore, the main control module calls the internally configured first-order hysteresis moving average filtering algorithm to smooth the transient analog signal collected by the sensing module, and has a built-in range dead zone verification mechanism. If any sensor feedback value is detected to exceed the reasonable physical range boundary, the effective value in the previous sampling period is used for subsequent calculations.

[0029] Furthermore, the static pressure sensor is connected to the fluid inlet pipe section of the first cavity with the pipe wall flush, and the pressure inlet of the dynamic pressure transmitter is connected to the fluid inlet and outlet of the first cavity respectively; the inlet temperature sensor is embedded in the fluid inlet pipe section of the first cavity.

[0030] Furthermore, the mechanical seal assembly includes a rotating ring component that rotates synchronously with the internal rotor shaft section and a stationary ring base fixedly arranged on the second cavity housing. The temperature probe of the thermocouple assembly is embedded inside the stationary ring base and close to the friction contact end face between the rotating ring component and the stationary ring base.

[0031] Furthermore, the heat exchange jacket is internally arranged with spiral guide ribs. The cooled gas undergoes convective and conductive heat transfer with the metal shell surface of the second cavity in the channel formed by the guide ribs, thereby enhancing the airflow turbulence and improving the heat transfer efficiency.

[0032] This invention provides an integrated multi-functional oil-gas mixed transportation system using a rotary pump. It offers the following advantages:

[0033] 1. This invention constructs a pneumatic unloading and cooling circulation structure by connecting a bypass pipeline between the fluid inlet and outlet of the first chamber and connecting a throttling and pressure-reducing valve in series in the bypass pipeline, in conjunction with a heat exchange jacket covering the outside of the second chamber. When dealing with high gas content or slug flow conditions, the system introduces the gas in the first chamber that would disrupt the liquid seal into the bypass pipeline for exhaust and unloading, preventing the loss of lubricating medium in the first chamber. At the same time, the gas generates an adiabatic expansion and cooling effect after passing through the throttling and pressure-reducing valve. The cooled gas enters the heat exchange jacket as a cold source to cool the oil in the second chamber, and the discharged waste gas flow is directly converted into a medium for cooling the mechanical seal components.

[0034] 2. This invention reads the absolute temperature of the mechanical seal assembly through the main control module, calculates the derivative of temperature with respect to time, and combines it with the quotient of the apparent gas flow rate and the critical suspension velocity of gas and liquid to generate a dynamic heat load index to adjust the oil circulation pump. This control logic integrates steady-state temperature value, transient temperature rise rate and aerodynamic disturbance feedforward, overcomes the physical hysteresis defect of traditional single temperature feedback control, and can respond in advance and adjust the oil circulation cooling intensity before the mechanical seal assembly experiences actual heat accumulation, thus avoiding the friction pair being in a dry friction state.

[0035] 3. The main control module of this invention has a built-in exponential adaptive adjustment function, which enables the target operating frequency of the oil circulation pump controlled by the variable frequency drive to be adaptively adjusted according to the dynamic heat load index. When the system heat load index is greater than the safety threshold, the oil circulation pump speed increases non-linearly to quickly provide a large flow of cold source; when the heat load index shows a downward trend, the speed is synchronously and slowly reduced, so that the system can allocate cooling intensity as needed according to transient operating conditions, avoiding long-term over-cooling of the oil medium and the energy consumption of the actuator. Attached Figure Description

[0036] Figure 1 This is a perspective view of the present invention;

[0037] Figure 2 This is a schematic diagram of the structure of the dual-cavity body portion of the present invention;

[0038] Figure 3 This is a schematic diagram of the internal cross-sectional structure of the dual-cavity body of the present invention;

[0039] Figure 4This is an architectural diagram of the rotor pump integrated multifunctional oil and gas mixed transportation system of the present invention;

[0040] Figure 5 The flowchart is a process for the integrated multifunctional oil-gas mixed transportation method of the rotor pump of the present invention.

[0041] Figure 6 This is the timing diagram of the parameter sensing and state initialization logic of the present invention;

[0042] Figure 7 This is a flowchart of the aerodynamic anti-entrapment prediction and venting control logic of the present invention;

[0043] Figure 8 This is a schematic diagram illustrating the preparation and coupling transfer principle of the trans-cavity thermodynamic cold source of the present invention;

[0044] Figure 9 This is a flowchart illustrating the logic of the multidimensional dynamic heat load advance assessment of the present invention.

[0045] Figure 10 This is the variable frequency adaptive cooling compensation and system closed-loop control logic diagram of the present invention;

[0046] Figure 11 This is a graph showing the aerodynamic disturbance characteristics and anti-entrapment early warning index of the present invention.

[0047] Figure 12 This is a comprehensive evaluation curve of the hysteresis thermodynamic temperature rise and cross-domain feedforward of the present invention.

[0048] Figure 13 The graphs show the traditional PID lag control and the frequency conversion adaptive advance compensation curves of this invention.

[0049] Among them, 10 is the dual-chamber body; 101 is the first chamber; 102 is the second chamber; 103 is the rotor body; 104 is the mechanical seal assembly; 20 is the pipeline module; 201 is the bypass pipeline; 202 is the throttling and pressure reducing valve; 203 is the heat exchange jacket; 30 is the sensing module; 301 is the static pressure sensor; 302 is the dynamic pressure transmitter; 303 is the thermocouple assembly; 304 is the inlet temperature sensor; 40 is the execution module; 401 is the servo regulating valve; 402 is the oil circulation pump; 403 is the frequency converter driver; and 50 is the main control module. Detailed Implementation

[0050] The technical solutions in the embodiments of the present invention will be clearly and completely described below with reference to the accompanying drawings. Obviously, the described embodiments are only some embodiments of the present invention, and not all embodiments. Based on the embodiments of the present invention, all other embodiments obtained by those skilled in the art without creative effort are within the scope of protection of the present invention.

[0051] See attached document Figure 1 - Appendix Figure 4 The present invention provides a rotor pump integrated multi-functional oil and gas mixed transportation system, which may include, at the hardware level, a dual-chamber body 10, a pipeline module 20, a sensing module 30, an execution module 40, and a main control module 50.

[0052] The dual-cavity body 10 is internally divided into a first cavity 101 and a second cavity 102 that are physically isolated. The first cavity 101 is provided with an internal rotor body 103, and the second cavity 102 is provided with a mechanical seal assembly 104 that is connected to the internal rotor body 103.

[0053] The piping module 20 includes a bypass line 201, a throttling and pressure reducing valve 202, and a heat exchange jacket 203. The bypass line 201 is connected between the fluid inlet and outlet of the first chamber 101. The throttling and pressure reducing valve 202 is connected in series in the bypass line 201. The heat exchange jacket 203 covers the outside of the second chamber 102. After passing through the throttling and pressure reducing valve 202, the bypass line 201 is connected to the fluid inlet end of the heat exchange jacket 203. The fluid outlet end of the heat exchange jacket 203 is connected to the fluid inlet end of the first chamber 101.

[0054] The sensing module 30 includes a static pressure sensor 301, a dynamic pressure transmitter 302, a thermocouple assembly 303, and an inlet temperature sensor 304. The static pressure sensor 301, the dynamic pressure transmitter 302, and the inlet temperature sensor 304 are assembled at the fluid inlet and outlet of the first cavity 101, and the thermocouple assembly 303 is installed inside the mechanical seal assembly 104 in the second cavity 102.

[0055] The execution module 40 includes a servo regulating valve 401, an oil circulation pump 402, and a frequency converter 403. The servo regulating valve 401 is connected in series in the bypass line 201 and is located between the fluid outlet end of the first chamber 101 and the throttling and pressure reducing valve 202. The oil circulation pump 402 is arranged in the oil circulation loop of the second chamber 102. The frequency converter 403 is connected to the oil circulation pump 402, and the main control module 50 is connected to both the sensing module 30 and the execution module 40.

[0056] See attached document Figure 5 , Figure 5 This is a flowchart of a rotary pump integrated multifunctional oil-gas co-transport method according to an embodiment of the present invention; the present invention provides a rotary pump integrated multifunctional oil-gas co-transport method, including the following steps:

[0057] S10, the pressure and pressure difference values ​​inside the first cavity 101 are read through the sensing module 30, and the absolute temperature value of the mechanical seal assembly 104 is read.

[0058] S20, calculate the apparent gas velocity and critical gas-liquid suspension velocity in the first chamber 101 based on the pressure and pressure difference values, and when the quotient of the apparent gas velocity and critical gas-liquid suspension velocity reaches the trigger threshold, control the servo regulating valve 401 to open so as to guide the gas inside the first chamber 101 to the bypass pipeline 201.

[0059] S30, guide the gas in the bypass line 201 through the throttling and pressure reducing valve 202 for adiabatic expansion and cooling, and introduce the cooled gas into the heat exchange jacket 203 to exchange heat with the oil medium inside the second chamber 102. After absorbing heat, the gas flows out of the heat exchange jacket 203 and merges into the fluid inlet end of the first chamber 101.

[0060] S40 calculates the temperature derivative with respect to time based on the absolute temperature value, and generates a dynamic heat load index by calculating the quotient, absolute temperature, and temperature derivative with respect to time.

[0061] S50 generates a frequency adjustment signal based on the dynamic heat load index and sends it to the frequency converter 403 to adjust the speed of the oil circulation pump 402 so as to circulate and transport the cooled oil medium in the heat exchange jacket 203 to the surface of the mechanical seal assembly 104.

[0062] See attached document Figure 6 , Figure 6 This is a timing diagram of parameter sensing and state initialization logic according to an embodiment of the present invention; during system startup and continuous operation cycle, the main control module 50 performs the initial loading of basic physical property parameters and high-frequency sampling of on-site physical quantities. The working process includes the following steps.

[0063] The main control module 50 reads and loads the environmental physical property parameters of the fluid medium inside the dual-cavity body 10. The main control module 50 is equipped with a storage chip, which is pre-written with fluid physical property constants that match the current oilfield gathering and transportation conditions. According to the fluid analysis report of the current oil production block, the operator inputs the initial component parameters through the host computer interface. The parameters called and loaded by the main control module 50 specifically include the liquid phase density of the working fluid under standard conditions. The molar mass of the gas being transported and the surface tension of the gas-liquid two-phase interface The aforementioned static physical property constants are usually determined based on the on-site working conditions. For example, when the transported medium is conventional water-containing associated gas, the liquid phase density can be taken as 850 to 1050 kg / m³. 3 Empirical calibration values ​​within the interval.

[0064] The above parameters are used as state baseline variables in subsequent calculations of apparent gas velocity and assessments of hydrodynamic instability at the gas-liquid interface. In this embodiment, the process of transmitting environmental parameter data via the communication bus is set according to the standard settings of the field industrial control network protocol.

[0065] After loading the static reference parameters, the sensing module 30 continuously detects the aerodynamic flow field state and thermodynamic environment parameters inside the first cavity 101, and feeds back the measured analog signals to the main control module 50. The static pressure sensor 301 is connected to the fluid inlet pipe section of the first cavity 101 with a flush mounting method to collect the transient static pressure of the airflow entering the cavity. The pressure inlet of the dynamic pressure transmitter 302 is connected to the fluid inlet and outlet of the first chamber 101, respectively, to measure the inlet and outlet pressure difference generated when gas flows through the working section of the internal rotor body 103. The inlet temperature sensor 304 embedded in the fluid inlet pipe section of the first cavity 101 is used to capture the absolute temperature of the intake airflow in real time. .

[0066] After receiving the instrument signals, the main control module 50 uses the analog-to-digital conversion channel to parse the analog signals into digital sequences with timestamps. Multiphase flow transportation is often accompanied by gas-liquid pulse disturbances. The main control module 50 calls the internally configured first-order hysteresis moving average filtering algorithm to smooth the transient spikes. Meanwhile, the main control module 50 has a built-in range dead zone verification mechanism. If any sensor feedback value is detected to exceed the reasonable physical range boundary, the effective value of the previous sampling cycle is retained. The reasonable physical range boundary range of the static pressure sensor 301 is set to 0.1 to 10.0 MPa, the boundary range of the dynamic pressure transmitter 302 is 0 to 2.5 MPa, and the boundary range of the inlet temperature sensor is 250 to 450 K. The effective value is the normal acquisition value within the above reasonable range range cached in the memory of the main control module 50 in the previous operation cycle. For example, the effective static pressure value acquired and latched in the previous cycle is 2.5 MPa, thus avoiding the crash of subsequent operation modules caused by erroneous data. Based on the acquired transient static pressure and absolute temperature, the main control module 50 can macroscopically calculate the transient gas phase density in the first cavity 101 according to the ideal gas state theory. Its basic physical correlation model is as follows:

[0067] ;

[0068] In the formula, The transient gas phase density inside the first cavity 101; The transient static pressure at the air inlet; The molar mass of the gas being transported; This is a universal gas constant, and its empirical values ​​typically range from 8.150 to 8.450. ; It represents the absolute temperature of the intake airflow.

[0069] Thermocouple assembly 303 detects the thermodynamic state of mechanical seal assembly 104 inside second cavity 102 and transmits absolute temperature values ​​to main control module 50. The mechanical seal assembly 104 includes a dynamic ring component that rotates synchronously with the shaft segment of the internal rotor body 103 and a stationary ring base fixedly arranged on the housing of the second cavity 102.

[0070] The temperature probe of thermocouple assembly 303 is embedded inside the stationary ring base and close to the friction contact end face of the moving and stationary rings. The temperature probe senses the transient heating state of mechanical seal assembly 104 in the initial stage of gas phase dry friction by relying on the thermal conduction of the metal solid medium. Thermocouple assembly 303 converts the end face heat into a microvolt-level thermoelectric potential signal and transmits it to the temperature transmission interface of main control module 50. The output range of this thermoelectric potential signal is usually set to 0 to 50mV and transmitted to the temperature transmission interface of main control module 50. Main control module 50 converts the thermoelectric potential into an absolute temperature value. The working monitoring range of this absolute temperature value is usually set to 273.15 to 473.15 K. Within the same hardware interrupt control cycle, the main control module 50 synchronously latches the pressure, differential pressure and various temperature values, thereby constructing a multi-dimensional physical quantity data matrix that is fully aligned in the time dimension, providing underlying data support for subsequent cross-domain parameter fusion calculations.

[0071] See attached document Figure 7 , Figure 7 This is a flowchart of the aerodynamic anti-entrapment prediction and venting control logic according to an embodiment of the present invention; after obtaining the basic physical quantity data matrix, the system needs to make a forward-looking judgment on the risk of flow field instability and liquid seal loss, and convert the judgment result into actual physical intervention actions.

[0072] The main control module 50 obtains the current actual gas phase velocity by establishing an aerodynamic evolution model inside the cavity. Based on Bernoulli's equation in fluid mechanics and the pipe resistance characteristics, the main control module 50 calculates the apparent gas phase velocity of the airflow passing above the main liquid surface inside the first cavity 101 under the current operating conditions. The apparent gas phase velocity reflects the shear kinetic energy of the transient airflow on the liquid remaining in the cavity. Combining the inlet and outlet pressure difference and transient gas phase density obtained in the previous steps, the main control module 50 uses the built-in velocity calculation formula to obtain the apparent gas phase velocity. The specific calculation model is as follows:

[0073] ;

[0074] In the formula, The apparent velocity of the gas phase within the first cavity 101; The pressure difference between the fluid inlet and outlet of the first chamber 101; The transient gas phase density inside the first cavity 101; The overall resistance coefficient of the internal flow channel of the first cavity 101 is dimensionless. In this embodiment, the overall resistance coefficient of the internal flow channel is... The size depends on the geometry of the internal rotor body 103 and the spacing of the baffles inside the cavity. Engineers can perform pressure drop tests on a standard clean water to air two-phase flow test bench to calibrate this coefficient before the system leaves the factory. For conventional multiphase flow rotor pumps, The calibration value range is usually between 1.5 and 4.5.

[0075] After obtaining the actual flow velocity, the main control module 50 constructs a gas-liquid critical suspension velocity model. This model is based on the gas-liquid two-phase flow interface dynamics principle and calculates the critical velocity boundary when the gas kinetic energy just overcomes the constraints of droplet gravity and surface tension. The main control module 50 calls the environmental physical property parameters loaded during the initialization phase to calculate the gas-liquid critical suspension velocity:

[0076] ;

[0077] In the formula, This is the critical velocity for gas-liquid suspension. An empirical constant characterizing the flow resistance and geometry inside the first cavity 101, which is dimensionless; The surface tension at the interface between the gas and liquid phases; The liquid density of the working fluid; This represents the transient gas phase density inside the first cavity 101. (Empirical constant) The values ​​are usually obtained by fitting the critical liquid carrying capacity test data of the fluid prototype in the multiphase flow test loop, and the values ​​range from 2.5 to 3.2.

[0078] After calculating the above two flow velocity parameters, the system can perform risk assessment and intervention closed loop. Since a single absolute value of flow velocity is difficult to intuitively reflect the instability risk under varying operating conditions, the main control module 50 calculates the ratio between the apparent gas phase velocity and the critical gas-liquid suspension velocity, and outputs a dimensionless dynamic anti-entrainment safety factor. :

[0079] ;

[0080] In the formula, The dynamic anti-pinch safety factor is dimensionless. The apparent velocity of the gas phase within the first cavity 101; It is the critical velocity for gas-liquid suspension.

[0081] The main control module 50 determines the safety factor of the dynamic anti-pinch system. Has the set trigger threshold been reached? In this embodiment, the trigger threshold is set internally by the main control module 50. The range is 0.80 to 0.85, when the dynamic anti-pinch safety factor is... Less than the trigger threshold If the gas flow energy is insufficient to disrupt the liquid surface tension, the liquid level in the first chamber 101 remains stable, and the main control module 50 does not output any intervention commands.

[0082] When the dynamic anti-pinch safety factor Reaching or exceeding the trigger threshold When the airflow is in a critical surge state, the lubricating fluid in the first chamber 101 is about to be entrained and lost on a large scale. At this time, the main control module 50 immediately triggers the early anti-cavitation logic, outputting a continuous analog opening control signal to the servo regulating valve 401 in the execution module 40. The main control module 50 has a built-in positive correlation continuous mapping function, and the opening of the servo regulating valve 401 changes with the airflow. Value exceeds The difference increases proportionally. After receiving the signal, the electric positioner inside the servo regulating valve 401 drives the actuator to change the valve core position, thereby opening the bypass channel according to the corresponding flow area.

[0083] With the opening of the servo regulating valve 401, the high-pressure pure gas accumulated inside the first chamber 101 is forcibly guided into the bypass pipeline 201, reducing the actual gas flow rate across the main liquid surface of the internal rotor body 103 in the physical path. This forces the apparent gas velocity in the first chamber 101 to drop instantaneously, thereby increasing the safety factor. Once the temperature drops to a safe range, this pneumatic unloading mechanism suppresses the risk of dry friction from an aerodynamic perspective, maintaining the lubricating fluid seal inside the first chamber 101. The PID control loop for the positioning control of the servo regulating valve 401 and the internal electrical drive hardware can be configured according to conventional industrial process control theory in this embodiment.

[0084] See attached document Figure 8 , Figure 8 This is a schematic diagram of the preparation and coupling transfer principle of a cross-chamber thermodynamic cold source according to an embodiment of the present invention; after the aerodynamic flow field inside the first chamber 101 is discharged and decelerated by the servo regulating valve 401, the system immediately converts the discharged high-pressure airflow into a physical cold source for cooling the mechanical seal. This conversion process specifically includes the following continuously operating steps.

[0085] The high-pressure gas guided by the servo regulating valve 401 to the bypass line 201 is forced to flow through the throttling and pressure reducing valve 202 under the drive of the inherent pressure difference at both ends of the line. The throttling and pressure reducing valve 202 has a throat orifice plate structure with a reduced flow area. When the high-pressure gas passes through this narrowed channel quickly, it is in a physical state with no mechanical work output and no obvious heat exchange with the outside, thus a typical adiabatic throttling process occurs.

[0086] From a thermodynamic perspective, due to the existence of intermolecular forces in real gases, when they experience a significant pressure drop under normal temperature and high pressure conditions, the intermolecular distance is forcibly widened, and some of the internal energy is converted into potential energy to overcome molecular attraction. Macroscopically, this manifests as a decrease in fluid temperature. Based on this physical phenomenon, the main control module 50 incorporates a quantitative thermodynamic evaluation model to predict the available cold source quality.

[0087] ;

[0088] In the formula, The absolute temperature of the low-temperature gas after throttling and expansion; The absolute temperature of the high-pressure gas entering the bypass line 201 before throttling; The Joule-Thomson coefficient is the coefficient for the fluid medium. The static pressure of the high-pressure gas before throttling; This refers to the static gas pressure at the outlet of the throttling and pressure reducing valve 202.

[0089] As a preferred method, the Joule-Thomson coefficient It is not a fixed value, but is determined by the gas's physical properties and the current temperature and pressure range. In actual engineering configurations, the main control module 50 has a multi-dimensional gaseous property matrix table burned into its memory. The main control module 50 obtains the gaseous property value under the current operating conditions through table lookup and linear interpolation algorithms. The numerical value is simultaneously obtained by the main control module 50 from the aforementioned steps, which is the transient static pressure at the intake inlet. Pressure difference between inlet and outlet The sum of these values ​​is approximately taken as the static pressure of the high-pressure gas before throttling. The input value.

[0090] Similarly, the gas flowing out of the throttling and pressure reducing valve 202 will eventually flow into the fluid inlet of the first chamber 101, which is in a low-pressure state, via the bypass line 201. The pressure drop along the way between the two is relatively small, and the main control module 50 directly controls the transient static pressure at the inlet. Approximate as the static gas pressure at the outlet of the throttling and pressure reducing valve 202 The input values ​​are used in the calculation. In addition, the absolute temperature of the high-pressure gas before throttling is also considered. The main control module 50, based on the aforementioned measured absolute temperature of the intake airflow, And combined with the empirical coefficient of multiphase flow pump for variable compression adiabatic temperature rise, a forward extrapolation value is assigned, or in the case where the compression temperature rise effect is not significant, the absolute temperature is directly assigned. As Approximate substitution is used to maintain the coherence of the algorithm chain.

[0091] When the system delivers moist natural gas and the first chamber 101 maintains a large working pressure difference, the above-mentioned isenthalpic expansion process can generate a stable low-temperature gas flow. As for the multi-stage pressure reduction component structure and anti-icing blockage structure design inside the throttling and pressure reducing valve 202, this embodiment can make conventional selections according to the actual pressure drop requirements.

[0092] After the low-temperature gas is generated, the system guides it to the target heating area to implement cross-chamber thermodynamic coupling. The cooled gas flows out from the throttling and pressure reducing valve 202 and is introduced into the heat exchange jacket 203 through the bypass line 201. The heat exchange jacket 203 covers the outside of the metal shell of the second chamber 102. The low-temperature gas flows in the internal channel of the heat exchange jacket 203 and undergoes convection and conduction heat transfer with the oil medium stored inside through the shell wall of the second chamber 102. Due to the temperature gradient between the low-temperature gas and the oil medium, heat is transferred from the oil medium through the metal wall to the low-temperature gas flow, so that the oil medium obtains deep physical cooling. This cross-interface heat transfer mechanism converts the gas that was originally discharged to maintain the stability of the flow field into cold energy to suppress the heating of the friction pair, realizing active intervention of the internal thermal field of the equipment. Spiral guide ribs can be arranged inside the heat exchange jacket 203 to increase the turbulence and residence time of the gas flowing through the shell surface.

[0093] As the heat exchange process is completed, the system needs to perform closed-loop recovery of the reheated fluid to maintain global pressure balance. The gas that has absorbed heat from the oil and reheated flows out of the fluid outlet of the heat exchange jacket 203. This part of the gas continues to flow towards the low-pressure side along the end of the bypass pipeline 201 and finally merges into the fluid inlet of the first chamber 101. Since the fluid inlet of the first chamber 101 is connected to the suction side of the rotor pump, its transient static pressure is at a low level in the entire pipeline environment. The reheated gas after throttling and depressurization can be smoothly drawn into the chamber to rejoin the main transport cycle. This pipeline topology constructs a closed driving pressure differential loop, which obtains cooling capacity by using the gas's own expansion to do work, while avoiding the discharge of the working medium into the external environment and ensuring the material balance of the system.

[0094] See attached document Figure 9 , Figure 9 This is a flowchart illustrating the multi-dimensional dynamic heat load advance assessment logic according to an embodiment of the present invention. After completing the preparation of the physical cold source and the construction of the cross-domain transport path, the system needs to perform a precise quantitative assessment of the heating state of the mechanical seal assembly 104 in order to provide a decision basis for the variable frequency drive. This assessment process is executed by the microprocessor inside the main control module 50, and specifically includes the following sequential calculation steps.

[0095] In this embodiment, due to the inherent time lag in heat conduction in solid media, a single absolute temperature value is insufficient to accurately reflect the deteriorating trend of transient heat generation in the friction pair. To address this issue, the main control module 50 collects absolute temperature values ​​from the thermocouple assembly 303 at high frequency. Time-domain differentiation is performed to extract the thermal hysteresis characteristics of the mechanical seal assembly 104. Considering the discrete sampling characteristics of the industrial control system, the main control module 50 uses a backward difference algorithm instead of continuous integral to calculate the derivative of temperature with respect to time. The specific discrete calculation model is as follows:

[0096] ;

[0097] In the formula, The temperature derivative of mechanical seal assembly 104 with respect to time; This represents the absolute temperature value for the current sampling period. This represents the absolute temperature value from the previous sampling period. The time interval set for two consecutive temperature samplings in the system.

[0098] To ensure that differential operations can accurately capture transient temperature rise trends while avoiding high-frequency electrical noise interference, the sampling interval time... The value of this time constant needs to match the hardware interrupt cycle of the main control module 50. As a preferred approach, this time constant is typically set between 0.01s and 0.05s; the derivative of temperature with respect to time... It physically characterizes the acceleration of heat accumulation at the contact end face of a mechanical seal, providing a basic reference for predicting the thermal shock the system will face.

[0099] In addition to real-time monitoring of the thermodynamic state, the system further introduces a cross-physics field feedforward intervention mechanism to achieve proactive regulation. The main control module 50 calls the dynamic anti-entrapment safety factor calculated in the aforementioned steps. When gas accumulates inside the first cavity 101, the dynamic anti-entrapment safety factor is reduced. When the temperature rises, the working fluid level of the internal rotor body 103 decreases accordingly, thereby weakening the lubrication environment of the mechanical seal assembly 104. This typically increases the risk of subsequent dry friction heat generation in actual operating conditions. The main control module 50 uses this dynamic anti-clamping safety factor. Introducing aerodynamic disturbance feedforward parameters into the thermodynamic evaluation loop, this parameter fusion mechanism that spans aerodynamics and thermodynamics overcomes the defect of traditional temperature closed-loop control that always lags behind the heating physical process, and enables cross-domain prediction of dry friction risk.

[0100] After acquiring multi-dimensional parametric features, the main control module 50 needs to aggregate them into a single control variable for direct use by the actuator. The main control module 50 will then use the absolute temperature value... The derivative of temperature with respect to time and dynamic anti-pinch safety factor Dimensionless processing and weighted calculations are performed to generate dynamic heat load indices to guide cooling capacity allocation. As a preferred method, the main control module 50 incorporates an engineering weighted evaluation formula:

[0101] ;

[0102] In the formula, This is a dynamic heat load index, dimensionless. This represents the absolute temperature value of the mechanical seal assembly 104 in the current cycle; This is the reference temperature under normal operating conditions of the mechanical seal; The temperature derivative of mechanical seal assembly 104 with respect to time; This is a reference value for the maximum permissible rate of temperature rise. The dynamic anti-pinch safety factor is dimensionless. The steady-state weighting coefficient for absolute temperature is dimensionless. is the dynamic weighting coefficient of the temperature derivative, which is dimensionless; is the feedforward weighting coefficient for aerodynamic disturbances, which is dimensionless.

[0103] Among them, the reference temperature Reference value for maximum temperature rise The material heat resistance properties of the mechanical seal assembly 104 and the factory bench test data are pre-written into the memory of the main control module 50, and each weighting coefficient must meet the following requirements. The constraint conditions, as a specific parameter setting in this embodiment, include the steady-state weighting coefficient. The preferred range is Dynamic weighting coefficient The preferred range is Feedforward weighting coefficients The preferred range is During the system commissioning phase, the three weighting coefficients can be calibrated by using the on-site friction temperature rise characteristic curve. For example, under conditions of severe fluctuations in multiphase flow, the feedforward weighting coefficient for aerodynamic disturbances can be appropriately increased. To enhance the system's sensitivity to predicting air resistance conditions, the proportion of [specific parameters] is adjusted. For the microprocessor hardware configuration that performs discrete difference operations and weighted summation inside the main control module 50, this embodiment can use a conventional industrial field programmable logic controller or digital signal processor.

[0104] See attached document Figure 10 , Figure 10This is a variable frequency adaptive cooling compensation and system closed-loop control logic diagram according to an embodiment of the present invention; after obtaining the dynamic heat load index characterizing the system's heating trend, the system needs to convert it into control commands for physical execution components in order to achieve on-demand distribution of cooling energy and closed-loop stability of the overall operating conditions.

[0105] The main control module 50 receives the dynamic heat load index generated in the previous calculation cycle and generates a frequency adjustment signal based on the index, which is then sent to the frequency converter 403. Considering that there is usually a nonlinear relationship between the temperature rise characteristics of the mechanical seal friction pair and the coolant flow rate, conventional linear proportional control will exhibit cooling compensation lag when the heat load suddenly increases. As a preferred approach, the main control module 50 internally constructs an exponential adaptive adjustment function to calculate the target operating frequency sent to the frequency converter 403. To avoid invalid calculations with negative bases caused by low-level fluctuations in the index, which could lead to software anomalies, this function uses piecewise constraint logic. The specific control logic formula is as follows:

[0106] ;

[0107] In the formula, The target operating frequency for the output of the frequency converter 403 to the oil circulation pump 402; Minimum operating frequency required to maintain basic lubrication of oil circulation pump 402 This is the gain coefficient for frequency control; This is a dynamic heat load index, dimensionless. The safe heat load threshold set for the system is dimensionless. It is a nonlinear response exponent, dimensionless.

[0108] The main control module 50 has pre-set specific boundary limits for the above parameters. In engineering practice, the minimum operating frequency is... It is usually set to a fixed value within the range of 15 to 20 Hz to prevent the oil pump from stalling at low frequencies and to ensure a safe thermal load threshold. As a control baseline for preventing dry friction, its preset value is generally calibrated to 1.0, and the nonlinear response index is... The preferred value range for this index is 1.2 to 1.5. This index allows the system to increase its frequency output with greater acceleration in the initial stage of excessive heat load. Furthermore, the gain coefficient for frequency control... The magnitude of the value reflects the system's response margin to thermal overload. During commissioning, it can be empirically calibrated based on the rated power of the oil circulation pump 402 and the pipe resistance characteristics of the cooling circuit. The value range is set between 5.0 and 15.0 Hz.

[0109] Upon receiving the frequency adjustment signal, the system actuator initiates physical cooling intervention; the frequency converter 403 adjusts the frequency according to the received target operating frequency. Adjusting the speed of the oil circulation pump 402, as the displacement of the oil circulation pump 402 increases, the oil medium that was originally stored inside the second chamber 102 and cooled by the low-temperature gas in the heat exchange jacket 203 is circulated and transported in a large flow rate and directly washes the metal surface of the mechanical seal assembly 104. Relying on the forced convection heat transfer effect, the oil medium quickly strips away the heat accumulated on the friction pair end area. The intervention of this composite cold source not only suppresses the rise in absolute temperature, but also alleviates the growth trend of the temperature derivative with respect to time, reducing the probability of thermal damage to the mechanical seal.

[0110] As cooling compensation continues and the aerodynamic flow field stabilizes, the system automatically enters the energy recovery stage. When the multiphase flow state in the first chamber 101 returns to stability and the heating of the mechanical seal assembly 104 is controlled, the dynamic heat load index calculated in real time by the main control module 50 is... The trend will show a decline. Based on the decay of the indicator, the main control module 50 will synchronously lower the target operating frequency through the aforementioned exponential adaptive adjustment function. This smoothly reduces the speed of the oil circulation pump 402. This dynamic optimization closed-loop feedback regulation mechanism avoids a sudden increase in oil viscosity and wasted pumping losses caused by prolonged excessive cooling, ensuring equipment operation safety while also considering the overall system's energy efficiency ratio.

[0111] For the AC-DC-AC inverter circuit topology inside the variable frequency drive 403, the generation of pulse width modulation waveforms, and the stator magnetic field drive mode of the oil circulation pump 402, this embodiment can use the conventional AC motor variable frequency speed control specifications for electrical wiring and parameter configuration. Thus, the main control module 50 completes the entire workflow from physical parameter perception to equipment risk assessment, and then to cross-domain cold source scheduling and variable frequency closed-loop control, realizing data linkage between various hardware modules and multi-dimensional protection of fluid machinery.

[0112] Application Examples:

[0113] To further aid in understanding the technical solution of this invention, the following explanation uses the actual operating conditions of a gas-liquid mixed transport pumping station in an oilfield as an example:

[0114] In this embodiment, the application scenario is a pressurized transportation node for water-bearing associated gas in an oilfield block. The transportation equipment adopts a dual-rotor multiphase flow pump. The fluid properties and basic system parameters loaded by the main control module are as follows: the liquid phase density of the transportation medium is set in the range of 950.0 to 990.0 kg / m³. 3 The surface tension was calibrated within the range of 0.060 to 0.072 N / m, and the inversion calibration range of the comprehensive resistance coefficient of the internal flow channel was 2.8. 3.5, the critical liquid carrying empirical constant is set between 2.5 and 3.0, and the weighting coefficients for the multidimensional heat load assessment are allocated as follows: steady-state weighting coefficient is set between 0.30 and 0.40, dynamic weighting coefficient is set between 0.25 and 0.35, and feedforward weighting coefficient is set between 0.30 and 0.40. During system operation, the sum of these three coefficients is dynamically constrained to be equal to 11. The safe heat load threshold for frequency conversion adaptive adjustment is set between 0.95 and 1.0, the gain coefficient is set between 8.0 and 12.0 Hz, and the minimum operating frequency is set between 18 and 22 Hz.

[0115] The experimental verification process introduced a conventional industrial control scheme as a control group. The control group adopted a closed-loop temperature control strategy, in which the system only collected the absolute temperature of the mechanical seal end face and directly output frequency commands to the oil circulation pump through a preset PID algorithm. The test condition was set as a sudden gas blockage event during long-cycle operation. In the actual test, the aforementioned parameters were all selected within the set range and operated with a set of benchmark values ​​to maintain benchmark consistency: in the initial 0 to 15 seconds of the test, the rotor pump inlet maintained a stable gas content; at the 20th second, a large volume of pure gas was injected into the inlet pipe to simulate slug flow impact and induce a drop in the liquid level in the first chamber; then at the 50th second, the normal gas-liquid ratio was restored. During the test, the data acquisition system simultaneously recorded the flow field and thermodynamic state characteristics. The obtained test data are shown in Table 1.

[0116] Table 1: Dynamic response test data of cross-domain anti-dry friction control under multiphase flow sudden change conditions

[0117] Sampling time (s) Apparent gas velocity (m / s) <![CDATA[Anti-entrainment safety factor ( 1)]]> Mechanical seal temperature (K) Temperature derivative with respect to time (K / s) Dynamic heat load index Control group oil pump frequency (Hz) The oil pump frequency (Hz) in this embodiment 0 4.12 0.61 335.2 0.05 0.72 20.0 20.0 10 4.25 0.63 335.8 0.06 0.74 20.0 20.0 20 8.94 1.35 336.5 0.07 1.18 20.5 21.6 30 9.05 1.38 345.1 0.86 1.62 23.2 25.8 40 6.82 1.02 358.3 1.32 1.54 32.7 27.5 50 4.51 0.68 362.4 0.41 1.25 38.6 23.4 60 4.18 0.62 354.2 -0.82 0.81 35.1 20.0

[0118] According to Table 1 and Figures 11 to 13 Therefore, the cross-domain physical parameter fusion and feedforward intervention mechanism of the present invention has a direct engineering control effect when dealing with unsteady disturbances in multiphase flow.

[0119] During the stable operating range of 0 to 10 seconds, the apparent gas flow rate is maintained at the normal baseline, and the anti-entrapment safety factor is below the set threshold range; both control schemes output the basic operating frequency to maintain the system's normal lubrication state.

[0120] At 20 seconds after the slug flow disturbance occurs, the apparent gas velocity increases to 8.94 m / s, and the anti-entrainment safety factor... The temperature reached 1.35; due to the hysteresis of solid thermal conduction, the mechanical seal end face temperature only rose to 336.5K, and the temperature derivative with respect to time changed by a small margin. The control group relied on absolute temperature feedback, and the oil pump frequency was only adjusted to 20.5Hz without intervention. In this embodiment, the scheme relies on the introduction of aerodynamic disturbance feedforward parameters, and the dynamic heat load index quickly rose to 1.18, exceeding the safety benchmark range. The drive inverter increased the oil pump frequency to 21.6Hz in advance in the early stage of heating, and started the early cooling compensation.

[0121] During the disturbance evolution phase, from 30 to 40 seconds, dry friction heat generation became apparent. Due to the lack of cooling compensation in the early stage, the control group's end-face temperature rose to 358.3K at the 40th second, with a temperature rise rate of 1.32K / s. At this time, the PID algorithm of the control group output a significant correction action, raising the frequency to 32.7Hz, demonstrating the characteristics of lag control. In this embodiment, the solution intervened at the 20th second. Under the effect of cold source injection from the thermodynamic cross-chamber refrigeration module, the rising trend of the end-face temperature was alleviated. The dynamic heat load index guided the oil pump frequency to rise gradually, with the maximum output maintained at around 27.5Hz, avoiding overload of the inverter due to sudden frequency increase and pressure shock in the pipeline network.

[0122] During the recovery phase from the initial working condition downturn, lasting 50 to 60 seconds, the airflow returned to normal. Due to the accumulated control deviation from the earlier integral action, the control group still had a frequency as high as 38.6 Hz at the 50th second, resulting in excessive cooling and no power consumption of the system. In this embodiment, the exponential adaptive adjustment function adaptively reduced the frequency output based on the decline of the dimensionless evaluation index. By the 60th second, the oil pump frequency of this embodiment had returned to the base value of 20.0 Hz, and the end face temperature showed a normal cooling trend.

[0123] The above test data shows that the control strategy of the present invention connects the signal levels between fluid mechanics and thermodynamics; by constructing an anti-entrainment safety index and integrating it into the heat load assessment model, the system initiates frequency conversion protection before the temperature parameter fluctuates significantly. This control mechanism reduces the physical delay error inherent in the temperature closed-loop control itself, and, in conjunction with the cross-domain cold source scheduling of exhaust gas throttling refrigeration, reduces the probability of mechanical thermal damage to fluid machinery caused by sudden changes in operating conditions.

[0124] Although embodiments of the invention have been shown and described, it will be understood by those skilled in the art that various changes, modifications, substitutions and alterations can be made to these embodiments without departing from the principles and spirit of the invention, the scope of which is defined by the appended claims and their equivalents.

Claims

1. A rotary pump integrated multi-functional oil-gas mixed transportation system, characterized in that, It includes a dual-cavity body (10), a pipeline module (20), a sensing module (30), an execution module (40), and a main control module (50); The dual-cavity body (10) is internally divided into a first cavity (101) and a second cavity (102); the first cavity (101) is provided with an internal rotor body (103), and the second cavity (102) is provided with a mechanical seal assembly (104) connected to the internal rotor body (103). The piping module (20) includes a bypass line (201) bridging the fluid inlet and outlet of the first cavity (101), a throttling and pressure reducing valve (202) connected in series in the bypass line (201), and a heat exchange jacket (203) covering the outside of the second cavity (102); the bypass line (201) passes through the throttling and pressure reducing valve (202) and is connected to the fluid inlet end of the heat exchange jacket (203), and the fluid outlet end of the heat exchange jacket (203) is connected to the fluid inlet end of the first cavity (101); The sensing module (30) includes a static pressure sensor (301), a dynamic pressure transmitter (302), an inlet temperature sensor (304) assembled in the first cavity (101), and a thermocouple assembly (303) assembled inside the mechanical seal assembly (104). The execution module (40) includes a servo regulating valve (401) connected in series in the bypass line (201), an oil circulation pump (402) arranged in the oil circulation loop of the second chamber (102), and a frequency converter (403) connected to the oil circulation pump (402). The main control module (50) is communicatively connected to the sensing module (30) and the execution module (40).

2. The integrated multi-functional oil-gas mixing system with a rotor pump according to claim 1, characterized in that, The main control module (50) is configured to perform the following working steps: The sensor module (30) reads the pressure and pressure difference values ​​inside the first cavity (101) and the absolute temperature value of the mechanical seal assembly (104). Based on the pressure and differential pressure values, calculate the apparent gas flow rate and critical gas-liquid suspension velocity in the first chamber (101), and when the quotient of the apparent gas flow rate and the critical gas-liquid suspension velocity reaches the trigger threshold, control the servo regulating valve (401) to open so as to guide the gas inside the first chamber (101) to the bypass pipeline (201). The gas in the bypass pipeline (201) is guided to undergo adiabatic expansion and cooling through the throttling and pressure reducing valve (202), and the cooled gas is introduced into the heat exchange jacket (203) to exchange heat with the oil medium inside the second chamber (102). After absorbing heat, the gas flows out of the heat exchange jacket (203) and merges into the fluid inlet end of the first chamber (101). The derivative of temperature with respect to time is calculated based on the absolute temperature value, and the quotient, the absolute temperature value, and the derivative of temperature with respect to time are combined to generate a dynamic heat load index. A frequency adjustment signal is generated based on the dynamic heat load index and sent to the variable frequency drive (403) to adjust the speed of the oil circulation pump (402) so as to circulate and transport the cooled oil medium in the heat exchange jacket (203) to the surface of the mechanical seal assembly (104).

3. The integrated multi-functional oil-gas mixing system with a rotor pump according to claim 2, characterized in that, The main control module (50) calculates the transient gas phase density based on the transient static pressure at the airflow inlet obtained by the static pressure sensor (301); The apparent velocity of the gas phase is equal to twice the pressure difference between the fluid inlet and outlet of the first cavity (101) obtained by the dynamic pressure transmitter (302), divided by the product of the comprehensive resistance coefficient of the internal flow channel of the first cavity (101) and the transient gas phase density, and then the square root is taken. The critical suspension velocity of the gas-liquid is equal to the product of the surface tension of the gas-liquid two-phase interface and the difference between the liquid phase density of the working liquid and the transient gas phase density, divided by the square of the transient gas phase density, and the resulting value is raised to the power of 0.25 and then multiplied by the empirical constant characterizing the internal geometry of the first cavity (101).

4. The integrated multi-functional oil-gas mixed transportation system with a rotor pump according to claim 2, characterized in that, The quotient of the apparent gas velocity and the critical gas-liquid suspension velocity is a dimensionless dynamic anti-entrapment safety factor. When the dynamic anti-entrapment safety factor exceeds the trigger threshold, the main control module (50) outputs an analog opening control signal to the servo regulating valve (401), and the opening of the servo regulating valve (401) increases proportionally to the difference between the dynamic anti-entrapment safety factor and the trigger threshold.

5. The integrated multi-functional oil-gas mixed transportation system with a rotor pump according to claim 2, characterized in that, The process by which the main control module (50) calculates the dynamic heat load index is to sum the first, second and third items; The first item is the ratio of the absolute temperature value to the set reference temperature multiplied by the set steady-state weighting coefficient; The second term is the ratio of the temperature derivative with respect to time to the reference value of the maximum allowable temperature rise rate multiplied by a set dynamic weighting coefficient; The third term is the quotient multiplied by the set feedforward weight coefficient.

6. The integrated multi-functional oil-gas mixing system with a rotor pump according to claim 5, characterized in that, The main control module (50) internally constructs an exponential adaptive adjustment function to calculate the target operating frequency corresponding to the frequency adjustment signal; When the dynamic heat load index is less than or equal to the set safe heat load threshold, the target operating frequency is equal to the minimum operating frequency required for the oil circulation pump (402) to maintain basic lubrication; When the dynamic heat load index is greater than the safe heat load threshold, the target operating frequency is equal to the minimum operating frequency, plus the product of the nonlinear response exponent of the difference between the dynamic heat load index and the safe heat load threshold and the frequency control gain coefficient.

7. The integrated multi-functional oil-gas mixing system with a rotor pump according to claim 1, characterized in that, The main control module (50) calls the first-order hysteresis moving average filtering algorithm to smooth the analog signal collected by the sensing module (30), and when it detects that the feedback value of any sensor exceeds the range boundary, it uses the effective value in the previous sampling period as the input value for the current operation.

8. The integrated multi-functional oil-gas mixed transportation system with a rotor pump according to claim 1, characterized in that, The static pressure sensor (301) is installed flush with the pipe wall into the fluid inlet pipe section of the first cavity (101), and the pressure inlet of the dynamic pressure transmitter (302) is connected to the fluid inlet and outlet of the first cavity (101) respectively; the inlet temperature sensor (304) is embedded in the fluid inlet pipe section of the first cavity (101).

9. The integrated multi-functional oil-gas mixing system with a rotor pump according to claim 1, characterized in that, The mechanical seal assembly (104) includes a rotating ring component that rotates synchronously with the shaft segment of the internal rotor body (103) and a stationary ring base fixedly arranged on the housing of the second cavity (102). The temperature probe of the thermocouple assembly (303) is embedded inside the stationary ring base and close to the friction contact end face between the rotating ring component and the stationary ring base.

10. The integrated multi-functional oil-gas mixing system with a rotor pump according to claim 1, characterized in that, The heat exchange jacket (203) is internally arranged with spiral guide ribs. The gas passing through the throttling and pressure reducing valve (202) undergoes convection and conduction heat exchange with the shell surface of the second cavity (102) in the channel formed by the flow through the guide ribs.

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