Variable frequency control method and system for a twin-screw compressor

Abstract

The present application relates to the technical field of double screw compressor control, in particular to a double screw compressor variable frequency control method and system. The method collects the compressor inlet temperature, pressure, body temperature and exhaust pressure by constructing adaptive dynamic adjustment logic, constructs the actual state of steam and thermal stability factor, and constructs the comprehensive thermal stability margin in combination with the body temperature acceleration; further integrates the exhaust pressure margin to form a comprehensive permission index. Based on the index and the inlet pressure change rate, the proportional coefficient of the PID controller is adaptively adjusted, and the output of the PID controller is nonlinearly scaled and hard-limited to realize dynamic fine adjustment of the main motor frequency. The present application solves the problems of adjustment oscillation and equipment damage, realizes dynamic matching of frequency regulation and real-time thermal state of the equipment, and improves the operation stability and safety.

Description

Technical Field

[0001] This invention relates to the field of twin-screw compressor control technology, specifically to a twin-screw compressor frequency conversion control method and system. Background Technology

[0002] Due to the different operating conditions and characteristics, especially the constantly changing inlet conditions, twin-screw compressors need to adapt accordingly. As twin-screw compressors are typical positive displacement compressors, it is necessary to adjust the speed and optimize the system control logic to adapt to the inlet operating state, thereby achieving synchronous regulation of the outlet state. The variable frequency control method and system control logic of twin-screw compressors are crucial and important, ensuring that the compressor's performance can meet the operating requirements.

[0003] The twin-screw compressor uses fixed PID parameters, adjusting the main motor's operating frequency based on the monitoring values ​​from temperature and pressure sensors during compressor operation; the frequency transition time for frequency rise and fall of the inverter is adjusted by a fixed frequency. In scenarios with continuously changing inlet conditions, factors such as temperature transmission intervals, data transmission efficiency, and lag in the response of auxiliary systems (liquid injection, oil circuit) can easily lead to frequency regulation oscillations, excessively rapid frequency rise causing rotor dry running or overheating damage, severely impacting the safety and energy efficiency stability of the equipment. Summary of the Invention

[0004] To address the technical problem of frequency regulation oscillation leading to equipment damage when using a fixed PID controller to regulate a twin-screw compressor, the present invention aims to provide a variable frequency control method and system for a twin-screw compressor. The specific technical solution adopted is as follows: This invention proposes a variable frequency control method for a twin-screw compressor, the method comprising: Collect the compressor inlet temperature and pressure, the compressor body temperature, and the discharge pressure; Based on the compressor inlet temperature and standard temperature, pressure and standard pressure, the actual state of the steam is determined; a thermal stability factor is constructed by combining the rate of change of the body temperature with the actual state; a comprehensive thermal stability margin is constructed by combining the acceleration of the body temperature change with the thermal stability factor; the exhaust pressure margin is calculated based on the exhaust pressure and the preset maximum exhaust pressure, and a comprehensive permissible index is determined by combining the comprehensive thermal stability margin. By combining the rate of change of compressor inlet pressure and the comprehensive allowable index, the proportional coefficient of the PID controller is adjusted; the basic adjustment amount output by the PID controller is nonlinearly scaled and hard-limited to dynamically adjust the operating frequency of the compressor.

[0005] Furthermore, the method for obtaining the actual state of the steam includes: Calculate the temperature deviation ratio between the steam inlet temperature and the standard temperature during compressor operation; calculate the pressure deviation ratio between the steam inlet pressure and the standard pressure during compressor operation; combine the temperature deviation ratio and the pressure deviation ratio to determine the actual state of the steam; wherein, both the temperature deviation ratio and the pressure deviation ratio are negatively correlated with the actual state.

[0006] Furthermore, the method for obtaining the thermal stability factor includes: Slide a time window on the time sequence corresponding to the body temperature, and obtain the rate of change of the body temperature at the current moment based on the difference between the mean body temperature within the time window corresponding to the current moment and the previous moment. The thermal stability factor is obtained by combining the rate of change of the body temperature, the preset sensitivity coefficient, and the actual state of the steam; wherein the rate of change of the body temperature is negatively correlated with the thermal stability factor, and the actual state of the steam is positively correlated with the thermal stability factor.

[0007] Furthermore, the method for obtaining the overall thermal stability margin includes: The acceleration of the computer body temperature change; when the acceleration is a positive acceleration greater than zero, the acceleration is multiplied by a preset compensation coefficient to obtain the thermal penalty deduction; otherwise, the thermal penalty deduction is zero; the thermal stability factor is subtracted from the thermal penalty deduction, and the lower limit of the subtraction result is truncated to obtain the comprehensive thermal stability margin.

[0008] Furthermore, the method for obtaining the exhaust pressure margin includes: Extract the latest value from the exhaust pressure and use it as the current exhaust pressure; obtain the exhaust pressure margin based on the ratio of the current exhaust pressure to the preset maximum exhaust pressure.

[0009] Furthermore, the method for obtaining the comprehensive licensing indicator includes: The comprehensive thermal stability margin and the exhaust pressure margin are weighted and fused to obtain a comprehensive safety assessment value; the comprehensive safety assessment value is then truncated to a lower limit to obtain the comprehensive permit index.

[0010] Furthermore, adjusting the proportional coefficient of the PID controller includes: The difference between the preset target exhaust pressure and the current exhaust pressure is obtained to obtain the current deviation; Slide a time window on the time sequence corresponding to the compressor inlet pressure, and obtain the rate of change of the compressor inlet pressure at the current moment based on the difference between the average value of the compressor inlet pressure in the time window corresponding to the current moment and the previous moment. By combining the preset fluctuation gain coefficient with the rate of change of the compressor inlet pressure, a fluctuation adjustment term is obtained; by combining the preset initial proportional coefficient, the comprehensive allowable index and the fluctuation adjustment term, the proportional coefficient of the PID controller is adjusted when the current deviation is positive.

[0011] Furthermore, the nonlinear scaling and hard limiting of the basic adjustment value output by the PID controller to dynamically adjust the operating frequency of the compressor includes: After adaptively adjusting the PID proportional coefficient in the PID controller, the basic adjustment amount output by the PID controller is obtained; the preset upper and lower limits of the adjustment range are extracted; the limiting function is called to forcibly truncate the basic adjustment amount within the safe range formed by the lower and upper limits of the range, thus obtaining the actual adjustment amount; the actual adjustment amount is superimposed on the current operating frequency to dynamically adjust the operating frequency of the compressor.

[0012] Furthermore, after dynamically adjusting the operating frequency of the compressor, the method further includes: The compressor inlet temperature and pressure, body temperature, and exhaust pressure are re-acquired; based on the re-acquired compressor inlet temperature and pressure, body temperature, and exhaust pressure, the comprehensive allowable index and the rate of change of the compressor inlet pressure are updated; based on the updated comprehensive allowable index and the rate of change, the proportional coefficient of the PID controller is adjusted again.

[0013] The present invention also proposes a variable frequency control system for a twin-screw compressor, including a memory, a processor, and a computer program stored in the memory and executable on the processor. When the processor executes the computer program, it implements the steps of any one of the twin-screw compressor variable frequency control methods.

[0014] The present invention has the following beneficial effects: This invention provides comprehensive status data for the entire adaptive frequency conversion control system by collecting real-time data, enabling control to move away from fixed-frequency control and be based on actual equipment operation. It calculates the actual state of the steam, providing an accurate initial assessment for subsequent adjustments. A thermal stability factor is constructed by combining the rate of change of the machine body temperature, directly reflecting the dynamic balance between heat generation and dissipation, enabling rapid early warning of abnormal temperature increases. A comprehensive thermal stability margin is constructed by introducing the acceleration of the machine body temperature, establishing a proactive penalty for accelerating temperature increases and suppressing the risk of thermal runaway earlier. The exhaust pressure margin is calculated to assess the space between the current pressure and the safe upper limit, preventing the system from operating under overpressure boundaries. A comprehensive permissible index is formed by integrating the comprehensive thermal stability margin and the exhaust pressure margin, unifying the safety quantification boundary and ensuring that each frequency adjustment command is generated within the system's safety boundary. The PID proportional coefficient is dynamically adjusted by combining the comprehensive permissible index and the pressure change rate, enabling adaptive control response and resolving adjustment oscillations caused by fixed parameters. Nonlinear scaling and hard limiting of the PID output restrict the frequency adjustment within the safe range determined by the physical response capability of the auxiliary system, preventing rotor dry running or thermal runaway caused by excessively rapid frequency increases. This invention achieves dynamic matching between frequency regulation and real-time thermal state of the equipment through the above-mentioned adaptive dynamic control logic, thereby improving the operating stability, safety and lifespan of the compressor under varying operating conditions, while also optimizing system energy efficiency. Attached Figure Description

[0015] To more clearly illustrate the technical solutions and advantages in the embodiments of the present invention or the prior art, the drawings used in the description of the embodiments or the prior art will be briefly introduced below. Obviously, the drawings described below are only some embodiments of the present invention. For those skilled in the art, other drawings can be obtained based on these drawings without creative effort.

[0016] Figure 1 This is a flowchart of a variable frequency control method for a twin-screw compressor provided in one embodiment of the present invention. Detailed Implementation

[0017] The following description, in conjunction with the accompanying drawings, details the specific scheme of the variable frequency control method and system for a twin-screw compressor provided by the present invention.

[0018] Please see Figure 1 The diagram illustrates a flowchart of a variable frequency control method for a twin-screw compressor according to an embodiment of the present invention. The method includes: Step S101: Collect the temperature and pressure at the compressor inlet, the machine body temperature, and the discharge pressure.

[0019] During operation, twin-screw compressors experience variations in steam temperature and content, leading to different compression levels required for different steam volumes. Using fixed-frequency frequency adjustment of the inverter's up-and-down transition time or a fixed proportional gain in a PID controller can result in heat waste across different batches. Therefore, to adapt to changes in inlet conditions, it's necessary to collect data on the compressor inlet temperature and pressure, the compressor body temperature, and the discharge pressure. The compressor inlet pressure and temperature directly reflect the real-time operating status; the compressor body temperature reflects the internal thermal equilibrium; and the discharge pressure is used to assess the system's pressure safety boundary. Collecting these data is fundamental to constructing the entire adaptive variable frequency control process, ensuring that subsequent frequency control is based on the equipment's real-time status.

[0020] In some possible implementations of this invention, the steam temperature and pressure at the compressor inlet, the body temperature, and the exhaust pressure are recorded based on timestamps to construct original time-series data sequences for each type of data. The original time-series data sequences are then processed using a moving average filtering method to eliminate random noise and short-term fluctuations in the data, thereby obtaining a smooth data sequence for each type of data.

[0021] Step S102: Based on the compressor inlet temperature and standard temperature, pressure and standard pressure, determine the actual state of the steam; combine the rate of change of the body temperature and the actual state to construct a thermal stability factor; combine the acceleration of the body temperature change and the thermal stability factor to construct a comprehensive thermal stability margin; based on the exhaust pressure and the preset maximum exhaust pressure, calculate the exhaust pressure margin, and combine the comprehensive thermal stability margin to determine the comprehensive permissible index.

[0022] During the operation of a twin-screw compressor, low-pressure saturated steam needs to be compressed to medium pressure for process use. This process requires ensuring that the amount of steam entering the compressor (i.e., pressure and temperature) meets the compression requirements to maintain a stable outlet pressure. For example, when the compressor operates at a normal frequency of 40Hz, its inlet pressure is 0.3MPa and its outlet pressure is 1.0MPa. If fluctuations occur in the upstream heat exchanger, causing the inlet pressure to drop sharply from 0.3MPa to 0.25MPa within a short period, such as 30 seconds, it means that the compression ratio increases. In this case, if the compressor's original frequency is maintained, the outlet pressure will drop accordingly, failing to meet the process requirements. The control system needs to respond dynamically, rapidly increasing the operating frequency from 40Hz to 60Hz to adapt to the changed compression requirements and ensure stable system operation.

[0023] For the above operating conditions, the actual state of the steam is determined based on the compressor inlet temperature and standard temperature and pressure. Specifically, during the operation of a twin-screw compressor, the steam temperature and content will vary, resulting in different degrees of compression required for different steam volumes. Therefore, determining the actual state of the steam based on the compressor inlet temperature and standard temperature and pressure provides a fundamental basis for subsequent evaluation.

[0024] Furthermore, a thermal stability factor is constructed by combining the rate of change of the compressor's body temperature with the actual state. The rate of change of the compressor's body temperature directly reflects the dynamic balance between the heat generation rate and the heat dissipation rate during the compression process. When the inlet steam state changes abruptly, such as a sudden pressure drop or excessively rapid frequency increase, the compression ratio increases, and the heat generation per unit time increases sharply. If the liquid injection cooling and heat dissipation cannot remove the heat in time, the body temperature will rise rapidly. Constructing a thermal stability factor more accurately reflects the system's thermal stability.

[0025] Furthermore, when there is a lag in the compressor monitoring data, the acceleration of the compressor body temperature change reflects the actual temperature fluctuation trend during compressor operation and the increasing trend of the temperature change rate. By combining the acceleration of the compressor body temperature change and the aforementioned thermal stability factor, a comprehensive thermal stability margin can be constructed, which can more accurately quantify the safety margin of the current thermodynamic state.

[0026] Finally, after establishing a comprehensive thermal stability margin reflecting the compressor's thermal state, to further mitigate the risk of pressure fluctuations, the control system calculates an exhaust pressure margin based on the exhaust pressure and the preset maximum exhaust pressure; this exhaust pressure margin represents the space between the exhaust pressure and the safe upper limit. Therefore, by combining the comprehensive thermal stability margin and the exhaust pressure margin, a comprehensive permissible index is determined. This index plays a crucial boundary constraint role, quantifying the internal thermal balance state of the compressor with the external pressure safety boundary, ensuring that every frequency adjustment is performed within the system's safety boundary. This avoids potential faults such as rotor dry running, increased vibration, or exhaust overpressure that might occur if the frequency is adjusted solely based on the thermal stability factor while ignoring the dynamic pressure characteristics.

[0027] Step S103: Combine the rate of change of compressor inlet pressure and the comprehensive allowable index to adjust the proportional coefficient of the PID controller; perform nonlinear scaling and hard limiting on the basic adjustment amount output by the PID controller to dynamically adjust the operating frequency of the compressor.

[0028] The proportional gain of the PID controller is adjusted by combining the rate of change of the compressor inlet pressure and the comprehensive allowable index. Specifically, the proportional gain of the PID controller is adaptively adjusted according to the comprehensive allowable index and the rate of change of the inlet pressure, ensuring a sensitive response when the operating conditions are stable and a gentle adjustment when the risk increases. This solves the frequency regulation oscillation problem that easily occurs when the parameters of a fixed PID controller are faced with continuously changing inlet operating conditions.

[0029] Furthermore, the basic adjustment value output by the PID controller is nonlinearly scaled and hard-limited. This operation forcibly truncates the basic adjustment value within a safe range. Its core purpose is to achieve dynamic and fine-grained adjustment of the main motor frequency, ensuring that every frequency adjustment is performed within the system's safe boundaries and preventing the frequency adjustment amplitude from exceeding the equipment's capacity.

[0030] Following steps S101 to S103, this embodiment of the invention provides a variable frequency control method for a twin-screw compressor. This method constructs adaptive dynamic control logic, real-time acquisition of compressor inlet temperature and pressure, compressor body temperature, and exhaust pressure. Based on the actual state of the steam, the rate of change of the compressor body temperature, and the acceleration of the compressor body temperature, a comprehensive thermal stability margin is constructed, and this is combined with the exhaust pressure margin to form a comprehensive allowable index. This index is embedded in a cascade PID controller structure, allowing the proportional coefficient to adaptively adjust with the comprehensive allowable index and the rate of change of the compressor inlet pressure. The basic adjustment amount output by the PID controller is nonlinearly scaled and hard-limited, achieving dynamic and precise adjustment of the compressor's main motor frequency. This method solves the problems of regulation oscillation, auxiliary system response lag, and rotor overheating caused by fixed PID controller parameters and fixed frequency adjustment time. It achieves dynamic matching between frequency adjustment and the real-time thermal state of the equipment, avoiding equipment damage caused by excessively rapid frequency increases, and significantly improving the compressor's operational stability, safety, and service life under varying operating conditions.

[0031] Preferably, in some possible implementations of the embodiments of the present invention, the method for obtaining the actual state of steam includes: calculating the temperature deviation ratio between the temperature at the steam inlet and the standard temperature when the compressor is running; calculating the pressure deviation ratio between the pressure at the steam inlet and the standard pressure when the compressor is running; and determining the actual state of steam by combining the temperature deviation ratio and the pressure deviation ratio; wherein the temperature deviation ratio and the pressure deviation ratio are both negatively correlated with the actual state.

[0032] As a concrete example, the formula for calculating the actual state of steam is expressed as: .in, This indicates the actual state of the steam; This indicates the standard inlet temperature of the compressor during operation (unit: Kelvin K). This indicates the standard inlet pressure of the compressor during operation; This indicates the temperature at the compressor inlet during operation (unit: K). This represents the pressure at the compressor inlet during operation; exp() represents an exponential function with the natural constant e as the base. This represents the temperature deviation ratio; This represents the pressure deviation ratio. The inlet pressure of the compressor during operation and the standard pressure are both measured using an absolute pressure gauge.

[0033] The larger the absolute value of the difference between the compressor inlet temperature and the standard temperature, the smaller the calculated function value; similarly, the larger the absolute value of the difference between the compressor inlet pressure and the standard pressure, the smaller the calculated function value. In other words, the smaller the function value representing the actual state of the steam, the further the initial steam parameters are from the standard state. Conversely, the smaller the absolute value of the difference between the compressor inlet temperature and pressure and their corresponding standard parameters, the larger the calculated function value, and the closer the actual state of the steam is to the standard state. The actual state of the steam ranges from (0,1], with values ​​closer to 1 indicating closer proximity to the standard state.

[0034] In a specific example of this invention, the standard inlet pressure of the compressor during operation is the rated inlet pressure of the compressor or the steady-state operating pressure required by the upstream process. This value can be obtained directly by reading the rated operating condition data on the compressor's nameplate or by consulting the design process guidelines for the entire MVR system.

[0035] In a specific example of this invention, the standard temperature setting method for the compressor inlet during operation is as follows: based on the determined standard pressure of the compressor inlet during operation, the corresponding saturated steam temperature is obtained by querying the standard steam thermodynamic property table; the specific value can be adjusted by the implementer according to the equipment.

[0036] Preferably, in some possible implementations of the embodiments of the present invention, the method for obtaining the thermal stability factor includes: sliding a fixed-length time window on the time sequence corresponding to the body temperature, calculating the difference between the mean body temperature within the time window corresponding to the current moment and the previous moment, and dividing the difference by the time interval between the two moments to obtain the rate of change of the body temperature at the current moment; that is, subtracting the mean body temperature within the time window corresponding to the previous moment from the mean body temperature within the time window corresponding to the current moment, dividing the result by the time interval between the two moments to obtain the rate of change of the body temperature at the current moment; combining the rate of change of the body temperature, a preset sensitivity coefficient, and the actual state of the steam to analyze and obtain the thermal stability factor; wherein, the rate of change of the body temperature is negatively correlated with the thermal stability factor; and the actual state of the steam is positively correlated with the thermal stability factor.

[0037] As a specific example, the fixed-length time window is set to 2 seconds; however, the implementer can adaptively adjust the specific value according to the required device response sensitivity.

[0038] As a concrete example, the formula for calculating the thermal stability factor is expressed as: .in: Indicates the thermal stability factor; Indicates the sensitivity coefficient; Indicates the rate of change of body temperature; This represents the actual state of steam; exp() represents an exponential function with the natural constant e as the base.

[0039] This represents the rate of temperature change of the compressor body during operation. The larger the value, the higher the compressor's temperature change efficiency at the current stage. This is the sensitivity coefficient, used to adjust how sensitive the system is to temperature changes. The function performs an exponential mapping on the results. When there is a sudden increase in load, insufficient lubrication, or lag in cooling, the rate of temperature change increases. A decrease in frequency indicates a need to slow down the frequency increase or increase liquid injection, directly reflecting the dynamic characteristics of the compressor's thermal balance. This is calculated by combining the current actual state of the steam with the rate of change of the compressor body temperature. The larger the value, the better the system's thermal stability; conversely, a smaller value indicates an abnormal temperature change in the compressor.

[0040] In one specific implementation of this invention, the preset sensitivity coefficient is obtained through factory testing or equipment calibration parameters and is set to 0.35; the reference value range is 0.2-0.5. The specific calculation method is as follows: obtain the maximum safe temperature change rate of the twin-screw compressor under rated lubrication and liquid injection cooling conditions; take the reciprocal of the maximum safe temperature change rate as the preset sensitivity coefficient. For example, when the actual temperature change rate reaches the maximum safe temperature change rate, the exponential term... This triggers significant safety degradation intervention; the maximum safe temperature change rate depends on the difference in thermal expansion coefficients between the twin-screw compressor rotor and the casing, as well as the design clearance. The specific value of the maximum safe temperature change rate is the maximum permissible transient temperature rise index recorded in the manufacturer's equipment nameplate safety manual, or it can be directly obtained through a bench test applying the maximum permissible speed and minimum liquid injection volume to the equipment's limit boundaries. For example, the maximum safe temperature change rate is typically between 2K / s and 5K / s, and this invention does not specifically limit it.

[0041] Preferably, in some possible implementations of the embodiments of the present invention, the method for obtaining the comprehensive thermal stability margin includes: calculating the acceleration of the temperature change of the computer body; when the acceleration is a positive acceleration greater than zero, multiplying the acceleration by a preset compensation coefficient to obtain a thermal penalty deduction, that is, using the product of the acceleration and the preset compensation coefficient as the thermal penalty deduction; otherwise, when the acceleration is less than or equal to zero, the thermal penalty deduction is zero; subtracting the thermal penalty deduction from the thermal stability factor, and truncating the lower limit of the subtraction result to obtain the comprehensive thermal stability margin. As a specific example, the formula for calculating the comprehensive thermal stability margin is expressed as: .in, Indicates the overall thermal stability margin; Indicates the thermal stability factor; The acceleration that indicates the change in body temperature; This represents the preset compensation coefficient; max() represents the maximum value function; The amount deducted for heat penalty.

[0042] The acceleration of the compressor's temperature change reflects the actual temperature fluctuation trend during operation, indicating the increasing trend of the temperature change rate. To intervene more proactively in the runaway temperature change trend, a penalty mechanism is introduced during the accelerated heating process. The `max()` function is used to filter out accelerations greater than zero, and a penalty is calculated only when the acceleration is greater than zero. The extracted acceleration is multiplied by a preset compensation coefficient to form the thermal penalty deduction for the thermal stability factor. The thermal penalty deduction is subtracted from the thermal stability factor to obtain the overall thermal stability margin. The `max()` function is used for lower limit truncation to prevent the overall thermal stability margin from being negative; the value range of the overall thermal stability margin is [0, 1].

[0043] In one specific implementation of this invention, the preset compensation coefficient is obtained through factory testing or equipment calibration parameters. The specific calculation method is as follows: obtain the maximum body temperature acceleration of the compressor within its design allowable range; set the preset compensation coefficient as the reciprocal of the maximum body temperature acceleration, that is, use the reciprocal of the maximum body temperature acceleration as the preset compensation coefficient. For example, when the actual body temperature acceleration reaches the maximum allowable value, the penalty term equals 1. If the thermal stability factor is close to 1 at this time, the penalty term can completely offset it, causing the overall thermal stability margin to rapidly decrease to 0, thereby achieving strong suppression of the thermal runaway trend. The maximum body temperature acceleration depends on the thermal shock resistance of the compressor's lubricating oil film and the thermal inertia of the rotor, and can be read by analyzing historical fault data during implementation. For example, for conventional industrial twin-screw compressors, the maximum body temperature acceleration is typically between 2K / s² and 10K / s², corresponding to a preset compensation coefficient range of 0.1 to 0.5, which is not specifically limited in this application.

[0044] It should be noted that the preset compensation coefficient is not a meaningless constant. Its physical meaning is the reciprocal of the maximum body temperature acceleration. Therefore, the unit of the preset compensation coefficient is s² / K. When the body temperature acceleration is multiplied by the preset compensation coefficient, a dimensionless thermal penalty reduction is obtained. This dimensionless thermal penalty reduction is subtracted from the dimensionless thermal stability factor to obtain the overall thermal stability margin, which is also a dimensionless value.

[0045] Preferably, in some possible implementations of the embodiments of the present invention, during compressor operation, gas compression causes a gradual increase in pressure and temperature, and sudden changes in operating conditions or frequency increase commands can cause a redistribution of the internal pressure and temperature fields. Therefore, it is necessary to avoid adjusting the frequency solely based on the thermal stability factor while ignoring the dynamic characteristics of pressure; the method for obtaining the exhaust pressure margin includes: The current exhaust pressure is obtained by extracting the latest value from the exhaust pressure. The exhaust pressure margin is then calculated based on the ratio of the current exhaust pressure to the preset maximum exhaust pressure. Specifically, the exhaust pressure margin is obtained by subtracting the ratio of the current exhaust pressure to the preset maximum exhaust pressure from 1. In other words, the difference between 1 and the ratio is used as the exhaust pressure margin. The exhaust pressure margin represents the margin between the exhaust pressure and the safe upper limit pressure. The larger the value, the safer the current exhaust pressure.

[0046] As a specific example, the preset maximum discharge pressure is the maximum discharge pressure allowed by the compressor design, which depends on the physical pressure resistance limit of the compressor body structure and the downstream process pipeline network; this value is obtained directly by reading the compressor's factory-calibrated maximum discharge pressure or analyzing historical damage data.

[0047] Preferably, in some possible implementations of the embodiments of the present invention, the method for obtaining the comprehensive permitting index includes: weighting and fusing the comprehensive thermal stability margin and the exhaust pressure margin to obtain a comprehensive safety assessment value; and truncating the comprehensive safety assessment value to obtain the comprehensive permitting index. As a specific example, the calculation formula for the comprehensive permitting index is expressed as: .in, Indicators of comprehensive licensing Indicates the weighting for temperature stability; This is expressed as the pressure margin weight; Indicates the preset license indicators; Indicates the overall thermal stability margin; Indicates the current exhaust pressure. This indicates the preset maximum exhaust pressure; max() represents the maximum value function.

[0048] Both temperature stability weight and pressure margin weight are set to coefficients greater than zero; the larger the overall thermal stability margin and exhaust pressure margin, the more stable the current system state, and the more sensitive the adjustment for the adaptive control system, i.e., the larger the overall allowable index; the max() maximum value function truncates the lower limit of the overall allowable index, updating the values ​​of the overall allowable index below the preset allowable index to the preset allowable index. During compressor operation, when both the overall thermal stability margin and the discharge pressure margin are zero, and the system is in an extremely unstable state, the existence of the comprehensive allowable index can prevent the control system from completely losing its adjustment capability, thereby enabling adaptive recovery.

[0049] As a specific example, the default license index is set to 0.3; however, the implementer can adjust this data according to the physical characteristics of the compressor hardware and the actual response capability of the auxiliary system.

[0050] As a specific example, the sum of the temperature stability weight and the pressure margin weight is set to 1. In this embodiment of the invention, the temperature stability weight and the pressure margin weight are used to balance the system's safety response bias to thermal anomalies and exhaust pressure anomalies. The specific values ​​of these two weights can be calculated by analyzing historical operational fault sample statistics of a specific model of compressor. Specifically, the proportion of faults caused by temperature anomalies in the historical data is used as the temperature stability weight; the proportion of faults caused by exhaust pressure anomalies is used as the pressure margin weight. Since the twin-screw steam compressor is a positive displacement machine with extremely small rotor clearance, its sensitivity to thermal expansion and temperature changes is significantly higher than its sensitivity to exhaust pressure. For example, a... =0.7, =0.3; The specific value can be reasonably determined by the implementer based on the actual operating conditions, safety requirements and historical data of the specific equipment, through methods such as equipment factory calibration or training with historical operating data.

[0051] Preferably, in some possible implementations of the embodiments of the present invention, adjusting the proportional coefficient of the PID controller includes: obtaining the difference between a preset target exhaust pressure and the current exhaust pressure to obtain the current deviation;

[0052] A fixed-length time window is slid across the time sequence corresponding to the compressor inlet pressure. The difference between the average compressor inlet pressure within the time window corresponding to the current moment and the previous moment is calculated. Specifically, the difference is obtained by subtracting the average compressor inlet pressure within the time window corresponding to the previous moment from the average compressor inlet pressure within the time window corresponding to the current moment. This difference is then divided by the time interval between the two moments to obtain the rate of change of the compressor inlet pressure at the current moment. A fluctuation adjustment term is obtained by combining a preset fluctuation gain coefficient with the rate of change of the compressor inlet pressure. The preset initial proportional coefficient, the comprehensive allowable index, and the fluctuation adjustment term are combined. When the current deviation is positive, the proportional coefficient of the PID controller is adjusted. The fixed-length time window is 2 seconds.

[0053] As a specific example, the preset target exhaust pressure can be set to 1.0 MPa; in actual industrial applications, the specific value is determined by the process requirements of the steam-using equipment downstream of the compressor, and the implementer can make adaptive adjustments as needed.

[0054] As a specific example, the fluctuation adjustment term is specifically represented as the product of a preset fluctuation gain coefficient and the rate of change of the compressor inlet pressure, plus a constant. The formula for calculating the proportional coefficient of the PID controller is expressed as: .in, This is expressed as the proportional coefficient of the PID controller; This represents the initial proportional coefficient of the PID controller; This is represented as a comprehensive licensing indicator; Represented as the fluctuation gain coefficient; Expressed as the rate of change of compressor inlet pressure; Indicates the fluctuation adjustment term; () represents the minimum value function; This indicates the preset maximum fluctuation amplification factor.

[0055] The initial proportional coefficient represents the normal response strength of the compressor in frequency regulation under standard steady-state conditions, providing a basic anchor point for the dynamic adjustment of the PID controller. It is a basic parameter that can be directly obtained before implementing this invention. The larger the comprehensive permissible index, the better the thermal stability of the machine body temperature and the more sufficient the exhaust pressure margin, allowing the PID controller to work with a more sensitive response and a larger frequency conversion amplitude. The larger the rate of change of the compressor inlet pressure, the stronger the external interference, which means that a stronger frequency conversion amplitude is required.

[0056] When the current deviation is positive, meaning the system needs to increase the frequency to raise the current exhaust pressure, if the system is in a stable thermal state and has sufficient pressure margin, a high proportional gain of the PID controller is maintained to ensure the agility of frequency adjustment. When the system is in a dangerous state of thermal anomaly or approaching the pressure limit, the upper limit of the minimum function is used to truncate large fluctuation adjustment terms, and the maximum value is the preset maximum fluctuation amplification factor. This significantly reduces the current proportional gain of the PID controller, forcing the controller's response to become sluggish, thereby avoiding equipment dry running or shutdown failures caused by excessive frequency conversion adjustment.

[0057] When the current deviation is less than or equal to zero, the system needs to reduce the frequency to lower the current exhaust pressure. At this time, the PID controller will not be subject to the attenuation limit of the comprehensive allowable index, and will directly maintain the preset initial proportional coefficient to control the frequency reduction, ensuring that the system can respond quickly and reduce the frequency at full speed to escape danger in times of crisis.

[0058] As a specific example, the preset maximum fluctuation amplification factor is set to 2.0; its specific value can be adjusted according to the wear and tear of the compressor equipment.

[0059] In one specific implementation of this invention, the fluctuation gain coefficient is set to 0.1; the value range is 0.05 to 0.2. Since the fluctuation gain coefficient is used to quantify the sensitivity of the PID proportional coefficient to the rate of change of inlet pressure, a smaller value is taken to prevent the control system from oscillating due to excessive compensation for pressure fluctuations. The specific value can be adjusted by the implementer based on the specific equipment and historical data training. It should be noted that the dimension of the fluctuation gain coefficient is s / MPa, and multiplying it by the rate of change of the compressor inlet yields a dimensionless value.

[0060] Preferably, in some possible implementations of the embodiments of the present invention, the basic adjustment amount output by the PID controller is nonlinearly scaled and hard-limited to dynamically adjust the operating frequency of the compressor, including: after adaptively adjusting the PID proportional coefficient in the PID controller, obtaining the basic adjustment amount output by the PID controller; extracting preset upper and lower limits of the adjustment range; calling the limiting function to forcibly truncate the basic adjustment amount within the safe range formed by the lower and upper limits of the range to obtain the actual adjustment amount; and superimposing the actual adjustment amount onto the current operating frequency to dynamically adjust the operating frequency of the compressor.

[0061] As a specific example, the formula for calculating the actual adjustment amount is expressed as follows: .in, Indicates the actual adjustment amount; Indicates the basic adjustment amount; Indicates the preset lower limit of the amplitude; Indicates the preset upper limit of amplitude; () represents the hard limiting function.

[0062] It should be noted that in the incremental PID control algorithm used in this embodiment of the invention, the difference between the current exhaust pressure and the preset target exhaust pressure is calculated as the current deviation. During normal operation of the compressor, the exhaust pressure directly determines the saturated steam temperature and the operating frequency of the compressor's main motor. The magnitude and trend of the current deviation directly determine the adjustment direction of the control system. The PID controller uses the current exhaust pressure and the standard pressure as inputs to calculate the current deviation, the cumulative amount of the current deviation, and the rate of change of the current deviation, which correspond to the proportional coefficient, integral coefficient, and derivative coefficient of the PID controller, respectively. In actual industrial control, in order to achieve frequency conversion control of the motor, the proportional coefficient, integral coefficient, and derivative coefficient set by the system all include the dimension conversion gain from pressure to frequency. In the incremental PID calculation, the time attribute of the sampling period is directly absorbed into the integral coefficient and derivative coefficient, so that after the current deviation, the cumulative amount of the current deviation, and the rate of change of the current deviation are multiplied by the above three coefficients, the dimensions of the calculated results of the proportional, integral, and derivative terms are uniformly aligned to the frequency dimension. Therefore, the basic adjustment amount output after the three terms are superimposed has a clear frequency dimension. The initial proportional gain set by the system is essentially calibrated in Hz / MPa. Therefore, after proportional calculation with its built-in dimension conversion property, the dimension of the basic adjustment quantity output by the PID controller naturally transforms into the frequency dimension. The incremental PID control algorithm described here is common knowledge in the field and will not be elaborated further.

[0063] In one specific implementation of this invention, Set to -3Hz; the reference range is -5Hz to -2Hz; the lower limit of the amplitude is set based on the steam pipeline network pressure stabilization requirements and anti-surge boundary to prevent excessive frequency reduction from causing inlet pressure buildup. The specific value can be adjusted by the implementer according to the specific parameters of the equipment and historical data.

[0064] In one specific implementation of this invention, Set to 2Hz; the reference range is 1Hz to 3Hz; the upper limit of the amplitude is set based on the response delay time of the injection system and the lubricating oil pump, which is used to prevent the rotor from dry running and instantaneous thermal runaway due to excessive frequency increase. The specific value can be adjusted by the implementer according to the specific parameters of the equipment and historical data.

[0065] Preferably, in some possible implementations of the embodiments of the present invention, after dynamically adjusting the operating frequency of the main motor, the method further includes: re-collecting the temperature and pressure of the compressor inlet, the body temperature, and the exhaust pressure; updating the comprehensive permitting index and the rate of change of the compressor inlet pressure based on the re-collected temperature and pressure of the compressor inlet, the body temperature, and the exhaust pressure; and adjusting the proportional coefficient of the PID controller again based on the updated comprehensive permitting index and the rate of change.

[0066] In summary, this invention constructs an adaptive dynamic control logic, collects real-time compressor inlet and internal temperatures and pressures, and builds a comprehensive thermal stability margin based on the actual steam state, the rate of change of compressor body temperature, and the acceleration of compressor body temperature. This is then integrated with the exhaust pressure margin to form a comprehensive allowable index. This method solves the problems of regulation oscillation, auxiliary system response lag, and rotor dry-running overheating caused by fixed PID parameters and fixed frequency adjustment time, achieving dynamic matching between frequency regulation and the real-time thermal state of the equipment. By embedding the comprehensive allowable index into the PID controller structure, the proportional coefficient is adaptively adjusted according to the rate of change of the comprehensive allowable index and inlet pressure. Furthermore, nonlinear scaling and hard limiting of the PID output prevent auxiliary system response lag and thermal anomalies caused by excessively rapid frequency increase. This significantly improves the compressor's operational stability, safety, and equipment lifespan under conditions of severe fluctuations in inlet operating conditions, while also optimizing system energy efficiency.

[0067] Based on the same inventive concept, the present invention also proposes a twin-screw compressor frequency conversion control system, including a memory, a processor, and a computer program stored in the memory and executable on the processor; when the processor executes the computer program, it implements the steps of any one of the twin-screw compressor frequency conversion control methods.

Claims

1. A variable frequency control method for a twin-screw compressor, characterized in that, The method includes: Collect the compressor inlet temperature and pressure, the compressor body temperature, and the discharge pressure; Based on the compressor inlet temperature and standard temperature, pressure and standard pressure, the actual state of the steam is determined; a thermal stability factor is constructed by combining the rate of change of the body temperature with the actual state; a comprehensive thermal stability margin is constructed by combining the acceleration of the body temperature change with the thermal stability factor; the exhaust pressure margin is calculated based on the exhaust pressure and the preset maximum exhaust pressure, and a comprehensive permissible index is determined by combining the comprehensive thermal stability margin. By combining the rate of change of compressor inlet pressure and the comprehensive allowable index, the proportional coefficient of the PID controller is adjusted; the basic adjustment amount output by the PID controller is nonlinearly scaled and hard-limited to dynamically adjust the operating frequency of the compressor.

2. The variable frequency control method for a twin-screw compressor according to claim 1, characterized in that, The method for obtaining the actual state of the steam includes: Calculate the temperature deviation ratio between the steam inlet temperature and the standard temperature during compressor operation; calculate the pressure deviation ratio between the steam inlet pressure and the standard pressure during compressor operation; combine the temperature deviation ratio and the pressure deviation ratio to determine the actual state of the steam; wherein, both the temperature deviation ratio and the pressure deviation ratio are negatively correlated with the actual state.

3. The frequency conversion control method for a twin-screw compressor according to claim 1, characterized in that, The method for obtaining the thermal stability factor includes: Slide a time window on the time sequence corresponding to the body temperature, and obtain the rate of change of the body temperature at the current moment based on the difference between the mean body temperature within the time window corresponding to the current moment and the previous moment. The thermal stability factor is obtained by combining the rate of change of the body temperature, the preset sensitivity coefficient, and the actual state of the steam; wherein the rate of change of the body temperature is negatively correlated with the thermal stability factor, and the actual state of the steam is positively correlated with the thermal stability factor.

4. The frequency conversion control method for a twin-screw compressor according to claim 1, characterized in that, The method for obtaining the overall thermal stability margin includes: The acceleration of the computer body temperature change; when the acceleration is a positive acceleration greater than zero, the acceleration is multiplied by a preset compensation coefficient to obtain the thermal penalty deduction; otherwise, the thermal penalty deduction is zero; the thermal stability factor is subtracted from the thermal penalty deduction, and the lower limit of the subtraction result is truncated to obtain the comprehensive thermal stability margin.

5. The frequency conversion control method for a twin-screw compressor according to claim 1, characterized in that, The method for obtaining the exhaust pressure margin includes: Extract the latest value from the exhaust pressure and use it as the current exhaust pressure; obtain the exhaust pressure margin based on the ratio of the current exhaust pressure to the preset maximum exhaust pressure.

6. The variable frequency control method for a twin-screw compressor according to claim 1, characterized in that, The method for obtaining the comprehensive licensing indicators includes: The comprehensive thermal stability margin and the exhaust pressure margin are weighted and fused to obtain a comprehensive safety assessment value; the comprehensive safety assessment value is then truncated to a lower limit to obtain the comprehensive permit index.

7. The variable frequency control method for a twin-screw compressor according to claim 1, characterized in that, The adjustment of the proportional coefficient of the PID controller includes: Obtain the difference between the preset target exhaust pressure and the current exhaust pressure to get the current deviation; Slide a time window on the time sequence corresponding to the compressor inlet pressure, and obtain the rate of change of the compressor inlet pressure at the current moment based on the difference between the average value of the compressor inlet pressure in the time window corresponding to the current moment and the previous moment. By combining the preset fluctuation gain coefficient with the rate of change of the compressor inlet pressure, a fluctuation adjustment term is obtained; by combining the preset initial proportional coefficient, the comprehensive allowable index and the fluctuation adjustment term, the proportional coefficient of the PID controller is adjusted when the current deviation is positive.

8. The frequency conversion control method for a twin-screw compressor according to claim 1, characterized in that, The nonlinear scaling and hard limiting of the basic adjustment value output by the PID controller to dynamically adjust the operating frequency of the compressor includes: After adaptively adjusting the proportional coefficient of the PID controller, the basic adjustment amount output by the PID controller is obtained; the preset upper and lower limits of the adjustment range are extracted; the limiting function is called to forcibly truncate the basic adjustment amount within the safe range formed by the lower and upper limits of the range, thus obtaining the actual adjustment amount; the actual adjustment amount is superimposed on the current operating frequency to dynamically adjust the operating frequency of the compressor.

9. The variable frequency control method for a twin-screw compressor according to claim 1, characterized in that, After dynamically adjusting the operating frequency of the compressor, the method further includes: The compressor inlet temperature and pressure, body temperature, and exhaust pressure are re-acquired; based on the re-acquired compressor inlet temperature and pressure, body temperature, and exhaust pressure, the comprehensive allowable index and the rate of change of the compressor inlet pressure are updated; based on the updated comprehensive allowable index and the rate of change, the proportional coefficient of the PID controller is adjusted again.

10. A variable frequency control system for a twin-screw compressor, comprising a memory, a processor, and a computer program stored in the memory and executable on the processor, characterized in that, When the processor executes the computer program, it implements the steps of the variable frequency control method for a twin-screw compressor as described in any one of claims 1 to 9.

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