Interstage pressure and temperature flexible control system of oil-free screw air compressor

Abstract

The invention relates to the technical field of air compressor regulation and control, particularly discloses an interstage pressure and temperature flexible control system of an oil-free screw air compressor, and aims to solve the problems that an existing system is fixed in interstage parameter, lagged in thermal inertia response, free of real-time control of compression efficiency, poor in load adaptability, prone to causing overtemperature potential safety hazards, insufficient in energy-saving potential and low in operation economy. The system comprises a first pneumatic adjusting module, a second pneumatic adjusting module, a temperature and pressure monitoring module, a PLC (Programmable Logic Controller) control module, a grading adjusting module and an opening adjusting module. Precise and flexible control over the pressure and temperature between the first stage and the second stage is achieved, then the problem that the minimum energy-saving rotating speed is limited due to the fact that the first-stage exhaust temperature is too high when the oil-free screw air compressor is loaded at the low rotating speed is solved, and the purposes that a unit operates safely at the lower rotating speed, and the energy-saving effect is further improved are achieved.

Description

TECHNICAL FIELD

[0001] The present application relates to the technical field of air compressor regulation, and particularly relates to a flexible control system for inter-stage pressure and temperature of an oil-free screw air compressor. BACKGROUND

[0002] The oil-free screw air compressor is widely used in chemical industry, electronics, medical treatment and other fields with high requirements for compressed air quality due to its characteristics of no oil pollution and stable operation. The inter-stage pressure and temperature between the first stage and the second stage of the oil-free screw air compressor directly determine the compression efficiency, equipment safety and energy saving effect, and are the core control objects.

[0003] The existing control of the inter-stage parameters of the oil-free screw air compressor has some defects. Firstly, fixed pipelines are used to connect the first stage, the second stage and the cooler, and the inter-stage compression ratio is a fixed value in the design stage, which cannot adapt to load fluctuations. When the load is low, the rotor speed is low, the gas leakage increases, the first stage exhaust temperature is easy to exceed the safety threshold, and the minimum speed of the unit energy saving is limited. When the load is high, the second stage inlet pressure is insufficient, the actual compression ratio of the second stage is large, and the second stage exhaust temperature is easy to exceed the safety threshold. Secondly, the temperature is passively reduced by relying on the intermediate cooler, and there is no active adjustment circuit. The response to rapid temperature changes lags, and temperature overshoots are easy to occur. Thirdly, the control logic is not associated with the compression efficiency. The temperature and pressure thresholds are set by experience, and the actual compression ratio and efficiency are not calculated in real time. When the efficiency is lower than 75%, the original parameters are still maintained for operation, and the energy consumption per unit exhaust volume increases by more than 13%.

[0004] Therefore, there is an urgent need for a flexible control system for inter-stage pressure and temperature of an oil-free screw air compressor to solve the above problems. SUMMARY

[0005] The present application aims to provide a flexible control system for inter-stage pressure and temperature of an oil-free screw air compressor, comprising: A first pneumatic adjustment module is used to add a first flexible pipeline between the outlet of the intermediate cooler and the first stage suction port, and a first pneumatic adjustment valve is installed in the middle of the first flexible pipeline. A second pneumatic adjustment module is used to add a second flexible pipeline between the second stage exhaust port and the inlet of the aftercooler, and a second pneumatic adjustment valve is installed in the middle of the second flexible pipeline. A temperature and pressure monitoring module is used to install a suction temperature sensor and a suction pressure sensor at the first stage suction port, install an exhaust temperature sensor and an exhaust pressure sensor at the first stage exhaust port, and install a total exhaust temperature sensor at the exhaust port of the whole machine. A PLC control module is used to receive sensor data of the temperature and pressure monitoring module in real time, and obtain a correction amount of the inter-stage compression ratio based on the difference between the actual exhaust temperature and the set temperature value and the deviation of the actual pressure from the design compression ratio. A hierarchical regulation module is configured to dynamically regulate the opening degrees of the first and second pneumatic regulating valves based on the inter-stage compression ratio correction amount under different load pressures. An opening degree regulation module is configured to generate valve opening degree instructions according to the inter-stage compression ratio correction amount, the first pneumatic regulating valve opening degree being calculated based on a negative exponential function of the correction amount, and the second pneumatic regulating valve opening degree being calculated based on a positive exponential function of the correction amount.

[0006] Further, the PLC control module comprises: A collection unit is configured to collect the load rate of the air compressor in real time, the load rate of the air compressor being calculated based on the current rotational speed and the rated rotational speed; A first calculation unit is configured to calculate a load correction coefficient, the load correction coefficient being determined based on the load rate in segments; A first correction unit is configured to correct the first pneumatic regulating valve opening degree, the corrected first pneumatic regulating valve opening degree being obtained by multiplying the initial opening degree by the load correction coefficient; A second correction unit is configured to correct the second pneumatic regulating valve opening degree, the corrected second pneumatic regulating valve opening degree being obtained by multiplying the initial opening degree by 2 and subtracting the difference of the load correction coefficient; A forced correction unit is configured to forcibly increase the corrected first pneumatic regulating valve opening degree by 20% to 35% when the actual exhaust temperature exceeds the alarm threshold.

[0007] Further, the temperature and pressure monitoring module comprises: A detection unit is configured to monitor the change rate of the primary exhaust temperature, the change rate of the primary exhaust temperature being calculated based on the temperature change value per unit time; A second calculation unit is configured to calculate a thermal inertia compensation coefficient, the thermal inertia compensation coefficient being determined based on the temperature change rate; An opening degree adjustment unit is configured to apply the thermal inertia compensation coefficient to the valve opening degree adjustment, the first pneumatic regulating valve opening degree being calculated by multiplying the corrected first pneumatic regulating valve opening degree by the compensation coefficient, and the second pneumatic regulating valve opening degree being calculated by multiplying the corrected second pneumatic regulating valve opening degree by 2 and subtracting the difference of the thermal inertia compensation coefficient; An emergency cooling unit is configured to trigger an emergency cooling mode when the temperature change rate is greater than a preset value.

[0008] Further, the PLC control module further comprises: A preset unit is configured to preset a primary exhaust pressure safety range, wherein the minimum safety pressure is 0.7 times the design pressure, and the maximum safety pressure is 1.3 times the design pressure; A first opening degree regulation unit is configured to adjust the first pneumatic regulating valve opening degree when the actual primary exhaust pressure is less than the minimum safety pressure; A second opening degree adjusting unit is configured to adjust the opening degree of the second pneumatic adjusting valve when the actual primary exhaust pressure is greater than the maximum safety pressure. A third opening degree adjusting unit is configured to close the second pneumatic adjusting valve and fully open the first pneumatic adjusting valve when the actual exhaust temperature exceeds 180 degrees Celsius.

[0009] Further, the hierarchical adjusting module comprises: A low load adjusting unit is configured to set the target compression ratio to 0.8 times the design compression ratio and raise the first pneumatic adjusting valve reference opening degree to the corresponding range in the low load mode. A medium load adjusting unit is configured to set the target compression ratio equal to the design compression ratio and enable closed-loop control to calculate the opening degree adjustment amount through the closed-loop control in the medium load mode. The integral calculation is obtained by multiplying the inter-stage compression ratio correction amount by a proportional coefficient and multiplying the integral of the inter-stage compression ratio correction amount by an integral coefficient. A high load adjusting unit is configured to set the target compression ratio to 1.2 times the design compression ratio and raise the second pneumatic adjusting valve reference opening degree to the corresponding range in the high load mode. A mode switching unit is configured to use a gradual transition method to limit the opening degree change rate to an increase or decrease of no more than 5% per second during mode switching.

[0010] Further, the temperature and pressure monitoring module further comprises: A third calculating unit is configured to calculate the change acceleration of the primary exhaust temperature, wherein the change acceleration is calculated by the change value of the temperature change rate per unit time. A generating unit is configured to generate a lead adjusting signal when the current change acceleration is greater than a preset maximum value of the change acceleration. A superimposing unit is configured to superimpose the lead adjusting signal to the first pneumatic adjusting valve opening degree instruction. A protection unit is configured to start an anti-overadjustment protection when the current change acceleration is less than a preset minimum value of the change acceleration.

[0011] Further, a calculating module is further included, and the calculating module comprises: A fourth calculating unit is configured to obtain the primary intake temperature and the primary intake pressure through the intake temperature sensor and the intake pressure sensor, obtain the primary exhaust temperature and the primary exhaust pressure through the exhaust temperature sensor and the exhaust pressure sensor, and calculate the actual compression ratio based on the primary intake temperature, the primary exhaust temperature, the primary intake pressure, and the primary exhaust pressure. A fifth calculating unit is configured to calculate the compression efficiency based on the actual compression ratio. A sixth calculating unit is configured to trigger an efficiency optimization mode when the compression efficiency is less than a preset efficiency value.

[0012] Furthermore, the present invention also discloses a computer device, including a memory and a processor, wherein the memory stores a computer program, and when the processor executes the computer program, it realizes the operation of a flexible control system of interstage pressure and temperature of an oil-free screw air compressor.

[0013] Furthermore, the present invention also discloses a computer-readable storage medium having a computer program stored thereon. When the computer program is executed by a processor, the operation of a flexible control system of interstage pressure and temperature of an oil-free screw air compressor is realized.

[0014] The beneficial effects of this application are: The present invention constructs a collaborative control architecture of flexible pipeline regulation loop, multi-measurement point monitoring, dynamic correction amount calculation and graded opening regulation, that is, a first pneumatic regulation module and a second pneumatic regulation module are set to construct flexible pipelines between the intercooler outlet and the first-stage intake port, the second-stage exhaust port and the aftercooler inlet respectively, and a temperature and pressure monitoring module is used to collect the temperature and pressure parameters of the first-stage intake, the first-stage exhaust and the whole machine exhaust. The PLC control module calculates the inter-stage compression ratio correction amount based on the temperature difference and the pressure deviation, and then adapts to different load pressure conditions through the graded regulation module. Finally, the opening regulation module regulates the opening of the two pneumatic regulating valves according to the exponential function law, so as to realize precise and flexible control of the pressure and temperature between the first and second stages, thereby solving the problem that the minimum energy-saving speed is limited due to the high exhaust temperature of the first stage when the oil-free screw air compressor is under low speed load, and achieving the goal of safe operation of the unit at a lower speed and further improving the energy-saving effect. BRIEF DESCRIPTION OF THE DRAWINGS

[0015] Figure 1 This is a schematic diagram of the system structure proposed in one embodiment of the present invention.

[0016] The realization of the objectives, functional features and advantages of this application will be further explained in conjunction with embodiments and with reference to the accompanying drawings. DETAILED DESCRIPTION

[0017] It should be understood that the specific embodiments described herein are only used to explain the present invention and are not intended to limit the present invention.

[0018] like Figure 1 As shown, the present application provides an interstage pressure and temperature flexible control system for an oil-free screw air compressor, comprising: The first pneumatic regulating module 1 is used to add a first flexible pipeline between the intermediate cooler outlet and the first-stage suction port, and a first pneumatic regulating valve is installed in the middle of the first flexible pipeline. The valve diameter of the first pneumatic regulating valve is determined based on the gas flow under the lowest load condition and the maximum allowable flow rate of the pipeline. The diameter value is calculated by multiplying the flow correction coefficient by the square root of the gas flow, dividing by the square root of the flow rate. The flow correction coefficient is between 0.8 and 1.2. The unit of the gas flow is cubic meters per second, and the flow rate is limited to no more than 25 meters per second. The second pneumatic regulating module 2 is used to add a second flexible pipeline between the second-stage exhaust port and the aftercooler inlet, and a second pneumatic regulating valve is installed in the middle of the second flexible pipeline. The valve diameter of the second pneumatic regulating valve is determined based on the diameter of the first pneumatic regulating valve. The diameter is calculated by multiplying the diameter of the first pneumatic regulating valve by a coefficient range of 0.7 to 1.1. The temperature and pressure monitoring module 3 is used to install a suction temperature sensor and a suction pressure sensor at the first-stage suction port, an exhaust temperature sensor and an exhaust pressure sensor at the first-stage exhaust port, and a total exhaust temperature sensor at the whole-machine exhaust port. The PLC control module 4 is used to receive the sensor data of the temperature and pressure monitoring module in real time, and to obtain a calculation inter-stage compression ratio correction value based on the difference between the actual exhaust temperature and the set temperature value and the deviation of the actual pressure from the design compression ratio. The specific steps include: first calculating the temperature difference, then calculating the pressure deviation value (by dividing the first-stage exhaust pressure by the first-stage suction pressure and then by the design compression ratio to obtain the deviation ratio), and finally multiplying the temperature difference by the temperature weight factor and the pressure deviation value by the pressure weight factor, and summing them up. The temperature weight factor is between 0.35 and 0.55, and the pressure weight factor is between 0.25 and 0.45. The step regulating module 5 is used to dynamically regulate the opening of the first pneumatic regulating valve and the second pneumatic regulating valve based on the inter-stage compression ratio correction value under different load pressures. The opening regulating module 6 is used to generate a valve opening instruction based on the inter-stage compression ratio correction value. The opening of the first pneumatic regulating valve is calculated based on the negative exponential function of the correction value, and the opening of the second pneumatic regulating valve is calculated based on the positive exponential function of the correction value. Specifically, the negative exponent of the correction value is multiplied by the valve characteristic constant and 100%. The opening of the second pneumatic regulating valve is calculated based on the positive exponential function of the correction value. Specifically, the negative exponent of the correction value is subtracted from 1, and then multiplied by the valve characteristic constant and 100%. The valve characteristic constant is between 0.8 and 1.5.

[0019] As described in the module 1-6 above, the present application adjusts the loop by constructing a flexible pipeline, monitors multiple points, calculates dynamic correction amount, and controls the cooperative control architecture of hierarchical opening degree, that is, the first and second pneumatic adjustment modules are respectively constructed in the flexible pipeline between the intermediate cooler outlet and the first suction port, and the second exhaust port and the rear cooler inlet, the temperature and pressure parameters of the first suction, the first exhaust, and the whole machine exhaust are collected by the temperature and pressure monitoring module, the inter-stage compression ratio correction amount is calculated by the PLC control module based on the temperature difference and the pressure deviation, and the hierarchical adjustment module is adapted to different load pressure conditions, and finally the opening degree adjustment module adjusts the opening degree of the two pneumatic adjustment valves according to the exponential function law, so as to realize the precise flexible control of the first and second inter-stage pressure and temperature, and solve the problem that the minimum energy-saving speed is limited due to the too high first exhaust temperature at low speed load of the oil-free screw air compressor, so as to achieve the goal of safe operation of the unit at a lower speed and further improve the energy-saving effect.

[0020] The compression process of the oil-free screw air compressor depends on the meshing of the head rotor to realize the volume compression of the gas, and the compression efficiency is positively correlated with the speed. When the unit is in a low-speed load condition, the decrease of the head rotor speed will increase the gas leakage in the rotor meshing gap, and the compression efficiency will decrease. The gas cannot fully complete the volume compression in the first compression process, and more input mechanical energy is converted into gas internal energy, which directly manifests as the increase of the first exhaust temperature. At the same time, the traditional oil-free screw air compressor connects the first and second head with fixed pipelines, and the first and second compression ratios are fixed values determined in the design stage, which cannot adapt to the characteristics of the decrease of the gas flow speed and the change of the pipeline pressure loss at low speed, and only relies on the passive cooling function of the intermediate cooler, which is difficult to effectively control the first exhaust temperature. If the speed is continued to be reduced for the purpose of energy saving, the first exhaust temperature will exceed the limit value of the head sealing element and the rotor gap, which may cause head failure, thereby limiting the energy-saving minimum speed of the unit and failing to fully develop the energy-saving potential of low-speed operation.

[0021] The traditional inter-stage pressure and temperature control scheme of oil-free screw air compressor has three core deficiencies: first, the pipeline design lacks flexibility, and only passive cooling of the first-stage exhaust gas is achieved through an intercooler, without an active adjustment loop, which cannot change the intake and exhaust flow ratio of the first-stage compressor head, making it difficult to actively adjust the inter-stage compression ratio; second, the temperature and pressure monitoring dimensions are not comprehensive, and only the exhaust temperature and pressure parameters of the entire machine are monitored, lacking real-time monitoring of key measurement points such as first-stage intake (affecting the initial conditions of first-stage compression) and first-stage exhaust (directly reflecting the first-stage compression effect), resulting in one-sided control basis and inability to accurately judge the dynamic changes of inter-stage working conditions; third, the control logic lacks precision, and simple on-off control is often used, such as "starting the cooling fan when the temperature exceeds the threshold" and "stopping when the pressure exceeds the threshold", without calculating the inter-stage compression ratio correction based on the dynamic changes of temperature difference and pressure deviation, and without adjusting the valve opening degree for different load pressures, making it difficult to adapt to the differences in working conditions under low, medium and high loads, and the control flexibility and adaptability are poor. To address the above problems, the present scheme adds flexible pipelines and pneumatic regulating valves to build active adjustment capability, supplements key measurement point monitoring to improve control basis, and combines dynamic correction calculation and hierarchical opening degree control of PLC to form a systematic solution covering the entire process, addressing the passive, one-sided and low-precision problems of traditional schemes.

[0022] The core function of the first pneumatic regulating module is to build a backflow adjustment loop for first-stage compressed air. A first flexible pipeline is added between the outlet of the intercooler (the temperature of the first-stage compressed air after cooling has dropped to 40-60°C) and the first-stage suction port, and a first pneumatic regulating valve is installed in the middle of the pipeline, to realize the function of introducing the cooled first-stage compressed air back to the first-stage suction port. The backflow of cold air can dilute the intake temperature of the first-stage suction port, while adjusting the intake flow of the first-stage compressor head, thereby actively changing the first-stage compression ratio. The valve diameter is calculated based on the gas flow under the lowest load condition and the maximum allowable flow rate of the pipeline, with the specific formula being: where D represents the valve diameter, k represents the flow correction coefficient, represents the gas flow under the lowest load condition, represents the maximum allowable flow rate of the pipeline; The physical meaning of the calculation logic is that the gas flow is minimum under the lowest load condition, and it is necessary to ensure that the gas flow rate in the pipeline does not exceed 25 meters per second at this time (to avoid excessive flow rate leading to excessive pipeline pressure loss, generating turbulent noise, or low flow rate leading to slow adjustment response); the flow correction coefficient of 0.8 to 1.2 is used to adapt to the flow characteristic differences of different gas media (such as dry compressed air, compressed air containing trace impurities), to ensure that the diameter calculation result takes into account safety and adjustment flexibility. The required "gas flow under the lowest load condition" is obtained based on the linear correspondence between the rated displacement of the air compressor and the speed, for example, the rated displacement is 10 m³ / s when the rated speed of the air compressor is 1500 r / min, and the gas flow is 3.33 m³ / s when the lowest load speed is 500 r / min (the lowest energy-saving speed of the traditional scheme), which is proportional to the speed; the "maximum allowable flow rate of the pipeline" is preset to 25 meters per second based on the pressure resistance of the pipeline material and the flow resistance characteristics of the gas. Taking specific parameters as an example, if the gas flow under the lowest load condition is 3.33 m³ / s, the flow correction coefficient is 1.0, and the flow rate is 20 meters per second (lower than the limit of 25 meters per second to leave a safety margin), then the first pneumatic regulating valve diameter = 1.0 x ≈1.0 x 1.825 ÷ 4.472 ≈ 0.408 meters, which can ensure smooth gas flow in the pipeline under the lowest load, meeting the backflow adjustment requirement and not causing the first-stage suction pressure fluctuation to exceed the safety range of ± 0.02 MPa due to excessive flow rate.

[0023] The second pneumatic regulating module forms a cooperative regulating relationship with the first pneumatic regulating module. The function of supplementing the high-pressure gas of the secondary exhaust into the primary exhaust section is realized by adding a second flexible pipeline between the secondary exhaust port (high-pressure gas after secondary compression, the pressure is usually 0.8-1.2 MPa) and the inlet of the aftercooler (i.e. the pipeline after the primary exhaust and before entering the aftercooler, the pressure is usually 0.3-0.5 MPa), and installing a second pneumatic regulating valve in the middle of the pipeline. The supplemented high-pressure gas can regulate the primary exhaust pressure, indirectly change the inlet pressure of the secondary head, and further cooperatively optimize the secondary compression ratio. The valve diameter is determined based on the first pneumatic regulating valve diameter, which is calculated by multiplying a coefficient of 0.7 to 1.1. The value of the coefficient is set according to the design compression ratio ratio of the primary and secondary heads: since the design compression ratio of the secondary head is usually higher than that of the primary (such as a primary design compression ratio of 3.0 and a secondary design compression ratio of 4.0), the secondary exhaust flow and the primary backflow flow need to maintain an appropriate ratio. The coefficient of 0.7 to 1.1 can ensure that the gas supplement amount of the second pipeline matches the flow amount of the first pipeline, avoiding the imbalance of inter-stage pressure caused by excessive flow in a single pipeline. Here, the "first pneumatic regulating valve diameter" is the value calculated in the first pneumatic regulating module (such as 0.408 meters). If the coefficient is 0.9 (adapted to the ratio of the primary and secondary design compression ratios of 3:4), then the second pneumatic regulating valve diameter = 0.408 x 0.9 ≈ 0.367 meters. This diameter can make the gas supplement amount of the secondary exhaust match the cold gas flow of the primary backflow at a ratio of 1:1.2, cooperatively regulate the primary exhaust pressure, and provide stable inlet conditions for the secondary head.

[0024] The temperature and pressure monitoring module is the sensing center of the whole control system. Real-time parameter acquisition is achieved by installing sensors at key measuring points: an air intake temperature sensor (monitoring the first-stage intake temperature, such as a real-time monitoring value of 37.5°C) and an air intake pressure sensor (monitoring the pressure, usually 0.09-0.11 MPa) are installed at the first-stage air intake port, which are used to obtain the initial gas temperature and pressure of the first-stage compression, providing a basis for calculating the actual compression ratio of the first stage; an exhaust temperature sensor (monitoring the gas temperature after the first-stage compression, directly reflecting the first-stage compression efficiency and heat accumulation) and an exhaust pressure sensor (monitoring the gas pressure after the first-stage compression) are installed at the first-stage exhaust port, which are used to obtain the core parameters of the first-stage compression effect; a total exhaust temperature sensor (monitoring the final output gas temperature, usually 40-50°C) is installed at the total exhaust port, which is used to assist in judging the overall operation state of the unit. The analog signals (such as DC 4-20 mA current signals) collected by all sensors are transmitted to the digital controller (such as the digital controller PIC-11, PIC-13 mentioned in the disclosure), which are converted into digital signals by the digital controller and then transmitted to the PLC control module, providing real-time and comprehensive parameter basis for subsequent correction calculation. For example, when the first-stage air intake temperature sensor collects Ts1 as 37.5°C and the first-stage exhaust temperature sensor collects a temperature of 135°C, the temperature rise of the first-stage compression can be directly obtained as 97.5°C, which provides a core basis for determining whether to increase the cold gas flow backflow and reduce the first-stage exhaust temperature.

[0025] The PLC control module undertakes the calculation center function, based on the real-time data transmitted by the temperature and pressure monitoring module, quantitatively calculates the inter-stage compression ratio correction amount, which is the core basis for subsequent regulation of the valve opening degree, and the specific calculation process is divided into three steps: the first step is to calculate the temperature difference value, that is, "temperature difference value = actual exhaust temperature - set temperature value", wherein "actual exhaust temperature" is the real-time value collected by the first-stage exhaust temperature sensor, and "set temperature value" is the first-stage exhaust safety temperature limit value preset based on the temperature resistance characteristics of the air compressor head, which may cause the rotor gap to shrink, resulting in rotor contact and other faults; the second step is to calculate the pressure deviation value, that is, "pressure deviation value = (first-stage exhaust pressure ÷ first-stage suction pressure) ÷ design compression ratio", wherein "first-stage exhaust pressure" and "first-stage suction pressure" are real-time values collected by the first-stage exhaust and first-stage suction pressure sensors, respectively, and "design compression ratio" is the rated compression ratio (such as 3.0) preset based on the structural parameters such as the number of teeth and the meshing gap of the first-stage head rotor, and the pressure deviation value reflects the deviation degree of the actual compression ratio from the design value; the third step is to calculate the inter-stage compression ratio correction amount, that is, "correction amount = temperature difference value × temperature weight factor + pressure deviation value × pressure weight factor", wherein the temperature weight factor is 0.35 to 0.55, and the pressure weight factor is 0.25 to 0.45, and the temperature weight factor is higher than the pressure weight factor, because the excessively high first-stage exhaust temperature directly threatens the safe operation of the unit, and the temperature needs to be controlled by adjusting the compression ratio in priority; the pressure weight factor is used to consider the stability of the inter-stage pressure, so as to avoid that the pressure deviation is too large to cause the head load fluctuation to exceed the safety range of ±5%. For example, if the actual first-stage exhaust temperature is 130°C (the set temperature value is 120°C, and the temperature difference value is 10°C), the temperature weight factor is 0.4; the first-stage exhaust pressure is 0.6 MPa, the first-stage suction pressure is 0.2 MPa (actual compression ratio = 0.6 ÷ 0.2 = 3.0), the design compression ratio is 3.0 (pressure deviation value = 3.0 ÷ 3.0 = 1.0, i.e. no deviation), and the pressure weight factor is 0.3; then the inter-stage compression ratio correction amount = 10 × 0.4 + (1.0-1.0) × 0.3 = 4.0, which indicates that the first-stage exhaust temperature needs to be reduced by about 10°C by increasing the first-stage cold gas flow backflow and adjusting the compression ratio, so as to restore to the safe range.

[0026] The function of the staged regulation module is to realize the working condition adaptation. Based on the working condition characteristics under different load pressures, the opening degrees of the two pneumatic regulating valves are differentially regulated and controlled. The "different load pressures" are determined by the exhaust pressure of the whole machine collected by the exhaust pressure sensor. The specific division standard is: the exhaust pressure of the whole machine ≤ 0.6 MPa corresponds to low load (rotating speed ≤ 800 r / min), 0.6 MPa < exhaust pressure of the whole machine ≤ 0.9 MPa corresponds to medium load (rotating speed 800-1200 r / min), and the exhaust pressure of the whole machine > 0.9 MPa corresponds to high load (rotating speed > 1200 r / min). The "inter-stage compression ratio correction amount" is the calculation result of the PLC control module. The significance of the module lies in that: under the low load working condition, the compression efficiency of the first stage head is the lowest, and more cold gas flow reflux is needed to reduce the first stage exhaust temperature by mainly increasing the cold gas flow reflux through the first pneumatic regulating valve; under the medium load working condition, the compression efficiency and the gas flow characteristics are relatively stable, and the two valves need to be balanced to maintain the stability of the inter-stage parameters; under the high load working condition, the compression load of the second stage head increases, and the air supply amount needs to be optimized through the second pneumatic regulating valve to stabilize the second stage compression ratio. For example, under the low load pressure (exhaust pressure of the whole machine 0.5 MPa), the inter-stage compression ratio correction amount is 5.0, the staged regulation module preferentially regulates and controls the first pneumatic regulating valve to increase the opening degree, and the second pneumatic regulating valve maintains a small opening degree; under the high load pressure (exhaust pressure of the whole machine 1.0 MPa), the correction amount is 2.0, the module preferentially regulates and controls the second pneumatic regulating valve to increase the opening degree, and the first pneumatic regulating valve maintains the basic opening degree, so as to ensure that the regulation effect and the working condition demand are accurately matched.

[0027] The function of the opening degree regulation module is to convert the inter-stage compression ratio correction amount into specific valve action instructions, to realize the smooth regulation of the valve opening degree through the exponential function law, and to avoid the sudden change of the opening degree leading to the sharp fluctuation of the inter-stage pressure and temperature. Among them, the opening degree of the first pneumatic regulating valve is calculated based on the negative exponential function of the correction amount, and the formula is "first pneumatic regulating valve opening degree = negative exponential of correction amount × valve characteristic constant × 100%"; the opening degree of the second pneumatic regulating valve is calculated based on the positive exponential function of the correction amount, and the formula is "second pneumatic regulating valve opening degree = (1-negative exponential of correction amount) × valve characteristic constant × 100%"; the "valve characteristic constant" is valued at 0.8 to 1.5, which is determined based on the flow characteristic curve of the pneumatic regulating valve (such as 1.0 for a linear flow characteristic valve and 1.2 for an equal percentage flow characteristic valve), and is used to adapt the flow opening degree response characteristics of different valves. The significance of this calculation logic lies in that: the first pneumatic regulating valve controls the cold gas flow reflux amount, the larger the inter-stage compression ratio correction amount (the greater the temperature and pressure deviation), the more cold gas flow reflux is needed. The inter-stage compression ratio correction amount is represented by w, then the negative exponential function The opening degree gently rises when the correction amount increases (e.g., when the correction amount increases from 4.0 to 8.0, e^-4.0 is approximately 0.018, and e^-8.0 is approximately 0.0003, and the opening degree gently decreases from 1.8% to 0.03%, which needs to be combined with the reference opening degree adjustment of the step regulation module to ensure that the actual opening degree is consistent with the correction requirement), thereby avoiding the sudden rise of the opening degree and the sudden drop of the primary suction pressure; the second pneumatic regulating valve controls the secondary exhaust air supplement amount, and the larger the correction amount, the smaller the supplement amount, and the positive exponential function can realize the gentle decrease of the opening degree when the correction amount increases (e.g., when the correction amount increases from 4.0 to 8.0, the opening degree gently increases from 98.2% to 99.97%), thereby avoiding the sudden decrease of the opening degree and the sudden rise of the primary exhaust pressure. Taking 1.0 as an example of the valve characteristic constant, when the inter-stage compression ratio correction amount is 4.0, the opening degree of the first pneumatic regulating valve is e^-4.0 x 1.0 x 100% approximately 1.8%, and the opening degree of the second pneumatic regulating valve is (1-e^-4.0) x 1.0 x 100% approximately 98.2%, and the two form a reverse cooperative regulation, which together finely adjusts the primary compression ratio from 3.0 to 2.8, and realizes the gentle decrease of the primary exhaust temperature.

[0028] Through the cooperative action of the above six modules, the working condition changes of the oil-free screw air compressor under different rotating speed loads can be adapted in real time, the primary and secondary inter-stage pressures and temperatures are actively adjusted, the problem of the excessively high primary exhaust temperature at low rotating speed is effectively solved, the minimum energy-saving rotating speed of the unit can be lower than that of the traditional scheme, the primary exhaust temperature can be controlled below the limit threshold when the unit is running at low rotating speed, and the energy-saving effect of the unit is improved. The passivity and blindness of the traditional scheme are avoided, the control precision and flexibility are significantly improved, and the unit can stably adapt to the full working condition range with a load rate of 20%-100%.

[0029] In an embodiment of the present application, the PLC control module comprises: a collection unit configured to collect the load rate of the air compressor in real time, wherein the load rate of the air compressor is obtained by calculating the current rotating speed and the rated rotating speed; a first calculation unit configured to calculate a load correction coefficient, wherein the load correction coefficient is determined based on the load rate in sections; when the load rate is less than 40%, the coefficient is equal to 1.2 minus 0.3 times the load rate; when the load rate is between 40% and 70%, the coefficient is fixed at 0.85; and when the load rate is greater than 70%, the coefficient is equal to 0.7 plus 0.2 times the load rate; a first correction unit configured to correct the opening degree of the first pneumatic regulating valve, wherein the corrected opening degree of the first pneumatic regulating valve is obtained by multiplying the initial opening degree by the load correction coefficient; a second correction unit configured to correct the opening degree of the second pneumatic regulating valve, wherein the corrected opening degree of the second pneumatic regulating valve is obtained by multiplying the initial opening degree by 2 and subtracting the difference of the load correction coefficient; A forced correction unit is configured to force the corrected first pneumatic regulating valve opening to increase by 20% to 35% when the actual exhaust temperature exceeds the alarm threshold.

[0030] As described above, by constructing a closed-loop control logic, i.e., the acquisition unit obtains the air compressor load rate based on the rotational speed, the first calculation unit determines the load correction coefficient according to the load rate interval, the first correction unit and the second correction unit respectively correct the initial opening of the first pneumatic regulating valve and the second pneumatic regulating valve by different calculation methods, and the forced correction unit triggers the opening forced increase when the actual exhaust temperature exceeds the alarm threshold, so as to realize the precise adaptation of the valve opening to different load rate conditions of the air compressor, solve the problem of insufficient opening correction accuracy of the basic regulation logic when the load rate dynamically changes, ensure that the unit can maintain the stability of the inter-stage pressure and temperature in the low, medium and high load rate intervals, further expand the load range of the unit safe operation, and improve the adaptability of the system to load fluctuations.

[0031] The load rate of the oil-free screw air compressor is directly related to the rotor speed of the head, and the speed determines the compression process characteristics of the gas in the head: when the load rate decreases, the rotor speed decreases, the gas leakage ratio in the rotor meshing gap increases, and the gas cannot be fully compressed in the primary compression process, more mechanical energy is converted into gas internal energy, resulting in an intensified rising trend of the primary exhaust temperature; at the same time, the gas flow speed in the pipeline decreases under low load rate, and if the initial opening of the first pneumatic regulating valve is only calculated based on the inter-stage compression ratio correction amount, it may not provide enough cold gas flow backflow to suppress the temperature rise. When the load rate increases, the rotor speed increases, the primary compression efficiency improves, the gas compression amount increases, the inlet gas amount and pressure of the secondary head increase synchronously, and if the second pneumatic regulating valve still uses fixed logic to correct the opening, it is easy to cause the secondary exhaust pressure to fluctuate beyond the design range (usually ±0.05MPa), increasing the load of the head. In addition, when the actual exhaust temperature exceeds the alarm threshold of the conventional regulation capacity, the conventional opening correction speed cannot meet the rapid cooling demand, and if not intervened in time, it will cause the rotor to expand, the rotor gap to shrink, and even cause the rotor to collide and jam, causing a shutdown failure. Therefore, it is necessary to correct the valve opening based on the load rate and set an over-temperature forced intervention mechanism to adapt to the compression characteristics under different loads and cope with extreme over-temperature conditions.

[0032] The core function of the acquisition unit is to obtain the load rate parameter reflecting the actual load state of the air compressor, which is calculated by "load rate = current speed ÷ rated speed x 100%", wherein the "current speed" is collected in real time by the speed sensor installed at the rotor shaft end of the air compressor head (the collection frequency is 1 time / second, the signal type is DC4-20mA, and the corresponding speed range is 0-30000r / min), and the "rated speed" is the rated running speed of the air compressor determined based on the material strength and compression efficiency of the head rotor in the design stage (for example, 1500r / min, which is a fixed design parameter stored in the parameter library of the PLC control module). The load rate directly reflects the actual running load of the head rotor. For example, if the current speed is 450r / min and the rated speed is 1500r / min, the load rate = 450 ÷ 1500 x 100% = 30%, indicating that the unit is in a low load running state, and the first pneumatic regulating valve opening needs to be adjusted to increase the cold gas flow backflow; if the current speed is 1200r / min and the rated speed is 1500r / min, the load rate = 1200 ÷ 1500 x 100% = 80%, indicating that the unit is in a high load running state, and the second pneumatic regulating valve opening needs to be optimized to stabilize the secondary air inlet pressure. The load rate data collected by the acquisition unit is transmitted to the first calculation unit in real time, providing a core basis for subsequent correction coefficient calculation, and its physical meaning is to convert the mechanical running parameter of the head rotor into a quantifiable load index, providing a calculation benchmark matched with the actual working condition for opening correction.

[0033] The function of the first calculation unit is to calculate the load correction coefficient based on the load rate interval segmentation. The design logic of the coefficient is directly related to the unit operation characteristics under different load rates. The specific segmentation rules and calculation methods are as follows: when the load rate is less than 40%, the linear calculation method of "load correction coefficient = 1.2-0.3x load rate" is adopted. In this interval, the unit is in a low load state, the rotor speed is low, the gas leakage amount is large, and the first stage exhaust temperature is easy to rise. A larger correction coefficient is needed to increase the opening degree of the first pneumatic regulating valve and increase the cold gas flow backflow amount. Therefore, the coefficient increases as the load rate decreases. For example, when the load rate is 30%, the coefficient = 1.2-0.3x0.3 = 1.11, and when the load rate is 20%, the coefficient = 1.2-0.3x0.2 = 1.14. The lower the load rate, the larger the correction coefficient, and the more significant the opening degree of the first pneumatic regulating valve is increased. When the load rate is between 40% and 70%, the load correction coefficient is fixed at 0.85. In this interval, the unit is in a medium load state, the rotor speed is moderate (usually 600-1050r / min), the gas leakage amount and compression efficiency are in a stable balance state, and the first stage exhaust temperature and second stage inlet pressure fluctuate less. The fixed correction coefficient can ensure the stability of the opening degree correction logic and avoid parameter fluctuations caused by frequent adjustments. When the load rate is greater than 70%, the linear calculation method of "load correction coefficient = 0.7+0.2x load rate" is adopted. In this interval, the unit is in a high load state, the rotor speed is high, and the gas compression amount is large. The second stage compressor inlet pressure rises synchronously with the load, and the second pneumatic regulating valve opening degree needs to be optimized to stabilize the second stage compression process. Therefore, the coefficient increases as the load rate increases. For example, when the load rate is 80%, the coefficient = 0.7+0.2x0.8 = 0.86, and when the load rate is 90%, the coefficient = 0.7+0.2x0.9 = 0.88. The higher the load rate, the larger the correction coefficient, and the more suitable the second pneumatic regulating valve opening degree correction is for the high load demand of the second stage. The load correction coefficients calculated by the unit are transmitted to the first correction unit and the second correction unit, respectively. The physical meaning is that through the segmented coefficient design, the opening degree correction logic is accurately matched with the core demand of the unit under low, medium and high load, avoiding the adjustment error caused by a single coefficient.

[0034] The function of the first correction unit is to correct the initial opening degree of the first pneumatic regulating valve, and the correction method is "first pneumatic regulating valve corrected opening degree = first pneumatic regulating valve initial opening degree x load correction coefficient", wherein the "first pneumatic regulating valve initial opening degree" is the opening degree value calculated by the opening degree adjusting module (calculated based on the negative exponential function of the inter-stage compression ratio correction amount), for example, the first pneumatic regulating valve initial opening degree is 20%, the load rate acquired by the acquisition unit is 30%, and the load correction coefficient calculated by the first calculation unit is 1.11, then the first pneumatic regulating valve corrected opening degree = 20% x 1.11 = 22.2%. The significance of this correction logic is that the core function of the first pneumatic regulating valve is to control the cold gas flow backflow from the intermediate cooler outlet to the first-stage suction port, and when the load is low, the load correction coefficient is greater than 1, which can increase the initial opening degree to increase the cold gas flow backflow, directly reduce the first-stage suction port intake temperature, reduce the internal energy conversion in the first-stage compression process, and thus inhibit the increase of the first-stage exhaust temperature; when the load is medium, the coefficient is 0.85 (less than 1), which can appropriately reduce the initial opening degree to avoid excessive backflow leading to excessively low first-stage intake pressure (lower than 0.08 MPa); when the load is high, the coefficient increases with the increase of the load rate, which can maintain or slightly increase the opening degree to balance the first-stage compression efficiency and temperature control demand. For example, when the load rate is 80%, the first pneumatic regulating valve initial opening degree is 15%, the load correction coefficient is 0.86, and the corrected opening degree = 15% x 0.86 = 12.9%, which can ensure the stability of the first-stage exhaust temperature while avoiding excessive backflow affecting the first-stage compression amount, and ensuring the second-stage engine head intake demand.

[0035] The function of the second correction unit is to correct the initial opening of the second pneumatic control valve. The correction method is "corrected opening of the second pneumatic control valve = initial opening of the second pneumatic control valve × (2-load correction coefficient)", where "initial opening of the second pneumatic control valve" is also the opening value calculated by the opening adjustment module (calculated based on the positive exponential function of the inter-stage compression ratio correction amount). The calculation logic of "2-load correction coefficient" is based on the functional complementarity design of the second pneumatic control valve and the first pneumatic control valve. At low load, the load correction coefficient is large, and the "2-load correction coefficient" is small, which can reduce the opening of the second pneumatic control valve, reduce the second-stage exhaust air supply volume, and avoid excessive first-stage exhaust pressure. At high load, the load correction coefficient is large, and the "2-load correction coefficient" is also large, which can increase the opening of the second pneumatic control valve and increase the air supply volume to stabilize the second-stage intake pressure. For example, if the calculated initial opening of the second pneumatic control valve is 30% and the load factor is 30%, the load correction factor is 1.11. "2-load correction factor" = 2-1.11 = 0.89, resulting in a corrected opening of 30% × 0.89 = 26.7%. This opening reduces the amount of air supply at low loads, preventing the first-stage exhaust pressure from exceeding the design range of 0.3-0.5 MPa. If the load factor is 80%, the load correction factor is 0.86. "2-load correction factor" = 2-0.86 = 1.14, resulting in a corrected opening of 30% × 1.14 = 34.2%. This opening increases the amount of air supply at high loads, ensuring a stable second-stage intake pressure of 0.6-0.8 MPa, meeting the high compression requirements of the second stage. The physical significance of this unit lies in the fact that, through differentiated correction formulas, the opening of the second pneumatic control valve complements that of the first pneumatic control valve, jointly maintaining interstage pressure balance and avoiding parameter imbalance caused by single-valve adjustment.

[0036] The function of the forced correction unit is to cope with extreme over-temperature working conditions. When the actual exhaust temperature collected by the first-stage exhaust temperature sensor in the temperature and pressure monitoring module exceeds the alarm threshold (which is determined based on the maximum temperature resistance characteristics of the head seal, and the head exhaust temperature may reach 250-300°C or higher depending on the design of the head), the forced correction logic is triggered to increase the first pneumatic regulating valve's corrected opening by 20% to 35%. For example, if the first pneumatic regulating valve's corrected opening is 22.2% and the actual exhaust temperature reaches 165°C (exceeding the 160°C alarm threshold), it is forced to increase by 30%, so the final opening is 28.86%. The physical meaning of this unit is that when the actual exhaust temperature exceeds the adjustment capacity of the conventional correction logic, the first pneumatic regulating valve's opening is forced to increase, significantly increasing the cold gas flow backflow, achieving rapid cooling. The adjustment rate of the conventional correction logic is 1%-2% per second, while the forced correction can complete a 20%-35% opening increase in 1 second, shortening the cooling response time to 5-10 seconds, which is much faster than conventional adjustment, effectively avoiding the shutdown protection threshold of further rising of the actual exhaust temperature. For example, if the actual exhaust temperature starts from 165°C and the opening is forced to increase, the cold gas flow backflow increases by 30%, the first-stage air inlet temperature drops from 35°C to 30°C, and the first-stage exhaust temperature drops to 145°C within 8 seconds, returning to the controllable range of the conventional adjustment, avoiding production interruption caused by over-temperature shutdown of the unit.

[0037] Through the synergistic effect of the above-mentioned units, the load adaptation optimization of the basic control logic is achieved: in the low load range, the first pneumatic regulating valve's opening is increased by a larger load correction coefficient, effectively suppressing the rise of the first-stage exhaust temperature, allowing the unit to operate stably at 30% load rate (corresponding to a speed of 450 r / min), which is lower than the minimum stable load rate of the traditional scheme. In the medium load range, the fixed correction coefficient ensures parameter stability, with the first-stage exhaust temperature fluctuating within ±5°C and the second-stage air inlet pressure fluctuating within ±0.02 MPa; in the high load range, the second pneumatic regulating valve's opening is optimized, reducing the second-stage exhaust pressure overshoot; in the over-temperature working condition, the forced correction mechanism shortens the cooling response time, effectively avoiding over-temperature shutdown. Overall, through load rate dynamic adaptation and emergency intervention, the control accuracy and stability of the system in the full load range are further improved, providing more comprehensive protection for safe and energy-saving operation of the unit.

[0038] In an embodiment of the present application, the temperature and pressure monitoring module comprises: a detection unit for monitoring the change rate of the first-stage exhaust temperature, which is calculated based on the temperature change value per unit time and has a unit of degrees Celsius per second; The second calculation unit is configured to calculate a thermal inertia compensation coefficient, which is determined based on a temperature change rate, and the calculation process includes the following steps: first, dividing the temperature change rate by a reference change rate (1.5-3.0 degrees Celsius per second); then, taking the hyperbolic tangent function value; then, multiplying the compensation intensity factor (0.1-0.3); and finally, adding 1 to obtain the coefficient value. The opening adjustment unit is configured to apply the thermal inertia compensation coefficient to the valve opening adjustment, and the first pneumatic regulating valve opening is calculated by multiplying the first pneumatic regulating valve corrected opening by the compensation coefficient, and the second pneumatic regulating valve opening is calculated by multiplying the second pneumatic regulating valve corrected opening by 2 and then subtracting the difference of the thermal inertia compensation coefficient. The emergency cooling unit is configured to trigger an emergency cooling mode when the temperature change rate is greater than a preset value (5 degrees Celsius per second), and finally, the first pneumatic regulating valve opening is increased by 40%, and the second pneumatic regulating valve opening is reduced by 50%.

[0039] As described above, by obtaining the temperature change rate of the first-stage exhaust gas per unit time through the detection unit, the second calculation unit calculates the thermal inertia compensation coefficient based on the change rate in combination with the hyperbolic tangent function and the compensation intensity factor, and the opening adjustment unit applies the compensation coefficient to the corrected openings of the first and second pneumatic regulating valves to offset the influence of thermal inertia. The emergency cooling unit triggers extreme working condition intervention when the temperature change rate exceeds the preset value, thereby realizing the targeted solution to the two core problems of thermal inertia lag and rapid temperature loss of control in the inter-stage temperature control of the oil-free screw air compressor, ensuring the timeliness and stability of the adjustment response in the temperature dynamic change process, and avoiding the adjustment overshoot caused by thermal inertia or the temperature loss caused by rapid temperature change.

[0040] The inter-stage temperature regulation of oil-free screw air compressor has significant thermal inertia characteristics. The change of the first stage exhaust temperature depends not only on the cold gas flow backflow of the first pneumatic regulating valve, but also on the thermal capacity of the intermediate cooler, connecting pipeline, and machine head shell. When the first pneumatic regulating valve increases the opening to increase the cold gas flow backflow, the cold air needs to exchange heat with the inner wall of the pipeline and the heat exchange elements of the intermediate cooler, so as to gradually reduce the overall gas temperature in the pipeline. The change of the first stage exhaust temperature sensor reading exists time lag (usually 3-8 seconds), which is the regulation delay caused by thermal inertia. If the temperature regulation is based only on the instantaneous temperature value without considering the temperature change rate, it is easy to appear the regulation overshoot phenomenon. For example, when the first stage exhaust temperature rises at a rate of 2℃ / s from 120℃, the conventional regulation needs to wait for the temperature sensor reading to exceed the set value (such as 125℃) before starting to increase the valve opening. At this time, the thermal inertia has caused the actual gas temperature in the pipeline to continue to rise, and when the regulation takes effect, the temperature may have exceeded 130℃, forming an overshoot. In addition, when the unit encounters sudden working conditions (such as the load rate suddenly rises from 50% to 80%), the first stage exhaust temperature change rate will increase sharply (may exceed 5℃ / s), and the response speed of the conventional thermal inertia compensation regulation cannot meet the rapid cooling demand. If effective intervention measures are not taken in time, the temperature will break through the alarm threshold in a few seconds, threatening the safety of the equipment. Therefore, the thermal inertia compensation coefficient needs to be calculated by monitoring the temperature change rate to correct the valve opening in advance to offset the lag, and an emergency cooling mode is set to deal with the ultra-fast temperature change working condition.

[0041] The core function of the detection unit is to obtain the change rate of the first stage exhaust temperature in real time. This parameter is calculated by "temperature change rate=(current first stage exhaust temperature-previous first stage exhaust temperature)÷time interval". The "current first stage exhaust temperature" and "previous first stage exhaust temperature" are both from the exhaust temperature sensor installed at the first stage exhaust port in the temperature and pressure monitoring module (the acquisition frequency is set to 1 time / s to ensure that the dynamic change of the temperature can be captured in time). The "time interval" is the time difference between the two data acquisition times of the sensor (fixed at 1 second, which can ensure the real-time data and avoid increasing the data noise due to too short interval). The unit of the calculation result is degree Celsius per second. The temperature change rate directly reflects the dynamic change trend of the first stage exhaust temperature. For example, if the current first stage exhaust temperature is 122℃ and the previous temperature (1 second ago) is 120℃, then the temperature change rate=(122-120)÷1=2.0℃ / s, indicating that the temperature is slowly rising. If the current temperature is 125℃ and the previous temperature is 120℃, then the temperature change rate=5.0℃ / s, indicating that the temperature is rapidly rising. The temperature change rate data collected by the unit is transmitted to the second calculation unit in real time. Its significance is to convert the static instantaneous value of the temperature into a dynamic trend value, which provides the core basis for the subsequent calculation of the thermal inertia compensation coefficient, avoiding the regulation lag problem of the traditional regulation which only relies on the instantaneous value.

[0042] The function of the second calculation unit is to calculate the thermal inertia compensation coefficient based on the temperature change rate. The design logic of the thermal inertia compensation coefficient aims to dynamically match the compensation strength by quantifying the temperature change rate. The greater the temperature change rate, the more significant the adjustment lag caused by thermal inertia, and a larger compensation coefficient is required to enhance the opening adjustment strength. The smaller the temperature change rate, the weaker the thermal inertia effect, and a smaller compensation coefficient is required to avoid over-adjustment. The specific calculation process is divided into four steps: the first step is to calculate the ratio of the temperature change rate to the reference change rate, where the "reference change rate" is a preset benchmark value based on the temperature change characteristics of the oil-free screw air compressor under normal operating conditions (the value is 1.5 to 3.0 degrees Celsius per second, for example, set to 2.0 degrees Celsius per second, which corresponds to the temperature change rate under normal load fluctuations of the unit). This ratio is used to standardize the actual temperature change rate and eliminate the influence of absolute numerical differences under different operating conditions; the second step is to take the hyperbolic tangent function value (tanh function) of the above ratio. The characteristic of the tanh function is that when the input value is between -1 and 1, the output value changes linearly, and the input value exceeds When the output value is outside this range, it approaches ±1. This feature ensures that the compensation coefficient changes smoothly when the temperature change rate is near the reference change rate, avoiding sudden changes in the compensation coefficient due to small fluctuations in the temperature change rate. When the temperature change rate far exceeds the reference value, the compensation coefficient tends to a stable maximum value to prevent over-compensation. In the third step, the hyperbolic tangent function value is multiplied by the compensation intensity factor (ranging from 0.1 to 0.3, for example, set to 0.2. The compensation intensity factor determines the maximum adjustment range of the compensation coefficient to avoid excessive or weak compensation). In the fourth step, the above product is added by 1 to obtain the final thermal inertia compensation coefficient (the coefficient value is always greater than 1 to ensure that the compensation direction is to increase the opening adjustment strength). For example, when the temperature change rate is 2.0 degrees Celsius per second, the reference change rate is 2.0 degrees Celsius per second, and the compensation intensity factor is 0.2, the calculation process is: ratio = 2.0 ÷ 2.0 = 1.0, tanh (1.0) ≈ 0.7616, product = 0.7616 × 0.2 ≈ 0.1523, thermal inertia compensation coefficient = 0.1523 + 1 = 1.1523; if the temperature change rate drops to 1.0 degrees Celsius per second, the ratio = 0.5, tanh (0.5) ≈ 0.4621, product = 0.4621 × 0.2 ≈ 0.0924, compensation coefficient = 1.0924. It can be seen that when the temperature change rate decreases, the compensation coefficient decreases synchronously, achieving precise matching of compensation intensity and temperature change rate. The thermal inertia compensation coefficient calculated by this unit is transmitted to the opening adjustment unit. Its significance lies in converting the dynamic temperature change trend into a quantifiable basis for opening compensation, realizing smooth adjustment of the compensation intensity through mathematical functions, and avoiding the rigidity problem of traditional fixed compensation.

[0043] The function of the opening adjustment unit is to apply the thermal inertia compensation coefficient to the opening correction of the first and second pneumatic regulating valves, and to adopt differentiated correction logic for the functional differences of the two valves. The first pneumatic regulating valve controls the cold gas flow backflow amount, and the core function is to suppress the rise of the primary exhaust gas temperature, which needs to increase the opening with the increase of the compensation coefficient. Therefore, the calculation method of "first pneumatic regulating valve opening = first pneumatic regulating valve corrected opening x thermal inertia compensation coefficient" is adopted, wherein the "first pneumatic regulating valve corrected opening" is the opening value calculated by the first correction unit (which has been corrected in combination with the load rate). The second pneumatic regulating valve controls the secondary exhaust gas air charge amount, and the increase of the opening will indirectly increase the primary exhaust gas pressure. In the case of rapid temperature change, the opening needs to be appropriately reduced to avoid pressure-assisted temperature rise. Therefore, the calculation method of "second pneumatic regulating valve opening = second pneumatic regulating valve corrected opening x (2-thermal inertia compensation coefficient)" is adopted, wherein the "second pneumatic regulating valve corrected opening" is the opening value calculated by the second correction unit, and "2-thermal inertia compensation coefficient" ensures that the second pneumatic regulating valve opening correction coefficient decreases synchronously when the compensation coefficient increases (for example, when the compensation coefficient = 1.1523, 2-1.1523 = 0.8477; when the compensation coefficient = 1.0924, 2-1.0924 = 0.9076), realizing the reverse cooperative regulation of the two valves. For example, if the calculated first pneumatic regulating valve corrected opening is 22.2% and the second pneumatic regulating valve corrected opening is 26.7%, and the thermal inertia compensation coefficient is 1.1523, then the first pneumatic regulating valve opening = 22.2% x 1.1523 ≈ 25.6%, which is 3.4 percentage points higher than the corrected opening, increasing the cold gas flow backflow amount to offset the thermal inertia lag; the second pneumatic regulating valve opening = 26.7% x 0.8477 ≈ 22.6%, which is 4.1 percentage points lower than the corrected opening, reducing the air charge amount to avoid pressure-assisted temperature rise. The significance of this correction logic is that through the reverse cooperative regulation of the two valves, the active intervention of temperature control is enhanced, and the inter-stage pressure is basically stable (the pressure fluctuation can be controlled within ±0.02 MPa), avoiding the parameter imbalance caused by single valve regulation, such as the traditional first valve regulation, which can cause a pressure fluctuation of ±0.05 MPa. The pressure fluctuation amplitude is reduced by 56% through the synchronous correction of the second valve, realizing the cooperative stability of temperature and pressure.

[0044] The function of the emergency cooling unit is to cope with extremely rapid temperature change conditions. When the first-stage exhaust gas temperature change rate monitored by the detection unit is greater than 5 degrees Celsius per second (this preset value is determined based on the temperature limit of the machine head material and the sensor response speed, and a change rate of 5 degrees Celsius per second means that if no intervention is made, the temperature will rise from 120 degrees Celsius to 170 degrees Celsius in 10 seconds, exceeding the alarm threshold), the emergency cooling mode is triggered, and the opening degrees of the two valves are forcibly adjusted: the final opening degree of the first pneumatic regulating valve is increased by 40% based on the calculated value of the opening degree adjustment unit to maximize the cold gas flow return amount and achieve rapid cooling; and the final opening degree of the second pneumatic regulating valve is reduced by 50% based on the calculated value of the opening degree adjustment unit to minimize the amount of air supplement and avoid the pressure rising to exacerbate the temperature rise. For example, if the opening degree of the first pneumatic regulating valve calculated by the opening degree adjustment unit is 25.6% and the opening degree of the second pneumatic regulating valve is 22.6%, after the emergency cooling mode is triggered, the opening degree of the first pneumatic regulating valve = 25.6% x (1+40%) = 35.84%, the cold gas flow return amount is increased by 40% compared with the regular compensation adjustment, and the inlet temperature of the first-stage suction gas can be reduced from 35 degrees Celsius to 28 degrees Celsius in 3 seconds; the opening degree of the second pneumatic regulating valve = 22.6% x (1-50%) = 11.3%, the amount of air supplement is reduced by 50%, and the first-stage exhaust gas pressure is reduced from 0.4 MPa to 0.35 MPa, avoiding pressure-assisted temperature rise. When the temperature change rate exceeds the adjustment capacity of thermal inertia compensation, the extreme opening degree adjustment breaks the vicious cycle of temperature rise-thermal inertia lag-adjustment failure, shortens the cooling response time from 8-12 seconds in the regular compensation to 3-5 seconds, ensures effective control before the temperature exceeds the safety threshold, and triggers the emergency cooling mode. The first-stage exhaust gas temperature can be reduced from 130 degrees Celsius (change rate 5.5 degrees Celsius per second) to 120 degrees Celsius in only 4 seconds, effectively avoiding equipment shutdown due to over-temperature.

[0045] Through the synergistic effect of the above-mentioned units, fine control of the inter-stage temperature dynamic change is achieved: in the case of regular temperature change (change rate ≤ 3 degrees Celsius per second), the thermal inertia compensation coefficient can reduce the first-stage exhaust gas temperature adjustment overshoot and improve the adjustment stability; in the case of medium rapid change (3 < change rate ≤ 5 degrees Celsius per second), the compensation coefficient and valve correction can shorten the temperature response time to avoid the temperature approaching the alarm threshold; in the case of extremely rapid change (change rate > 5 degrees Celsius per second), the emergency cooling mode can suppress the temperature rise trend in 5 seconds, and the number of equipment over-temperature shutdowns is reduced compared with the traditional scheme. Overall, through temperature change rate monitoring and dynamic compensation, the adjustment lag and overshoot problems caused by thermal inertia are solved, and extreme temperature intervention is used to cope with extreme temperature changes, further improving the temperature control capability of the system under dynamic conditions and providing more comprehensive temperature protection for the safe and stable operation of the unit.

[0046] In an embodiment of the present application, the PLC control module further comprises: The preset unit is configured to preset a first-stage exhaust pressure safety interval, wherein a minimum safety pressure is 0.7 times a design pressure, and a maximum safety pressure is 1.3 times the design pressure. The first opening degree adjustment unit is configured to adjust the opening degree of the first pneumatic regulating valve when the actual first-stage exhaust pressure is less than the minimum safety pressure, and the opening degree is calculated by multiplying the final opening degree by a ratio of the actual pressure to the minimum safety pressure. The second opening degree adjustment unit is configured to adjust the opening degree of the second pneumatic regulating valve when the actual first-stage exhaust pressure is greater than the maximum safety pressure, and the opening degree is calculated by adding (the final opening degree minus 80%) to the final opening degree, multiplying the actual pressure minus the maximum safety pressure, and dividing by the maximum safety pressure. The third opening degree adjustment unit is configured to close the second pneumatic regulating valve and fully open the first pneumatic regulating valve when the actual exhaust temperature exceeds a set threshold.

[0047] As described above, by constructing a pressure closed-loop management system, i.e., by the preset unit explicitly defining the minimum and maximum safety boundaries of the first-stage exhaust pressure, the first opening degree adjustment unit adjusts the opening degree of the first pneumatic regulating valve when the pressure is lower than the minimum safety value, the second opening degree adjustment unit differentially adjusts the opening degree of the second pneumatic regulating valve when the pressure is higher than the maximum safety value, and the third opening degree adjustment unit triggers extreme valve action when the actual exhaust temperature exceeds the limit threshold, realizing full-condition safety control of the first-stage exhaust pressure, solving the problem of insufficient response accuracy of the above basic regulation logic when the pressure exceeds the conventional range, and insufficient intervention for extreme over-temperature conditions, ensuring that the first-stage exhaust pressure is always in a safe operating interval, while providing the last safety guarantee for over-temperature extreme conditions, further improving the safety and stability of system operation.

[0048] The first-stage exhaust pressure is a core parameter connecting the first-stage and second-stage compression processes of the oil-free screw air compressor, and its value directly determines the synergistic efficiency of the two-stage compression and equipment safety: when the first-stage exhaust pressure is lower than the minimum safety pressure, the intake pressure of the second-stage compressor head is insufficient, which will lead to a decrease in gas compression in the second-stage compression process, and the overall machine exhaust volume will be lower than the design value (usually reduced by 15%-25%), which cannot meet the downstream gas demand; when the first-stage exhaust pressure is higher than the maximum safety pressure, the compression load of the first-stage compressor head will exceed the design bearing range, the force at the rotor engagement will increase, which will easily lead to accelerated bearing wear, excessive vibration amplitude (exceeding 0.15 mm / s) of the compressor head, and long-term operation will shorten the service life of the compressor head; conventional opening degree regulation cannot quickly cool down, and extreme valve action is required to cut off the heating source and maximize the cooling effect to avoid irreversible failures such as melting of sealing elements and lubrication failure. Therefore, it is necessary to explicitly control the boundaries by presetting the first-stage exhaust pressure safety interval, design differentiating regulation logic for different pressure abnormal scenarios, and set an over-temperature extreme intervention mechanism to ensure the synergistic operation of two-stage compression and equipment safety.

[0049] The core function of the preset unit is to define the safe operating boundary of the first-stage exhaust pressure, that is, to preset the first-stage exhaust pressure safety range, where the "design pressure" is the first-stage exhaust rated pressure determined during the design phase of the oil-free screw air compressor's first-stage head based on structural parameters such as the number of rotor teeth, meshing clearance, and rated speed (for example, 0.4MPa, the first-stage exhaust rated pressure is a fixed design parameter, stored in the parameter library of the PLC control module, and directly determines the rated output of the first-stage compression and the design intake conditions of the second-stage head); "minimum safety pressure = design pressure × 0.7", the physical meaning of taking 0.7 times the design pressure is to ensure that the second-stage head obtains the minimum intake If the first-stage exhaust pressure falls below this value, the second-stage exhaust pressure will be insufficient, and the gas filling rate within the rotor meshing gap will fall below 80%, resulting in the second-stage compression volume failing to meet design requirements and a significant reduction in overall exhaust volume. The "maximum safety pressure = design pressure × 1.3" setting of 1.3 times the design pressure is intended to protect the first-stage exhaust from overload damage. If the first-stage exhaust pressure exceeds this value, the compression reaction force on the first-stage exhaust rotor will exceed the design load limit (usually 1.3 times the rated load), causing bearing wear to accelerate by 3-5 times, and the exhaust vibration amplitude to exceed the safety limit of 0.15 mm / s. Taking a design pressure of 0.4 MPa as an example, the preset first-stage exhaust pressure safety range is 0.28 MPa (0.4 × 0.7) to 0.52 MPa (0.4 × 1.3). This range ensures both second-stage exhaust volume and first-stage safety, providing a clear control benchmark for subsequent pressure adjustments.

[0050] The function of the first opening adjusting unit is to cope with the scenario that the primary exhaust pressure is lower than the minimum safe pressure, and to increase the pressure by adjusting the opening of the first pneumatic regulating valve. The adjustment logic is "first pneumatic regulating valve adjusted opening = first pneumatic regulating valve final opening x (actual primary exhaust pressure ÷ minimum safe pressure)", where "actual primary exhaust pressure" is obtained from the exhaust pressure sensor installed at the primary exhaust port in the temperature and pressure monitoring module (acquisition frequency 1 time / s, ensuring real-time capture of pressure dynamics), "minimum safe pressure" is a value determined by the preset unit (e.g. 0.28 MPa), and "first pneumatic regulating valve final opening" is the corrected opening value (which has been combined with load rate and thermal inertia compensation, e.g. 25.6%). The first pneumatic regulating valve controls the cold gas flow backflow from the intermediate cooler outlet to the primary suction port. When the primary exhaust pressure is lower than the minimum safe pressure, the cold gas flow backflow needs to be reduced to increase the primary suction port intake pressure, thereby increasing the exhaust pressure of the primary compression. Through the proportional coefficient of "actual pressure ÷ minimum safe pressure", the opening adjustment range and the pressure loss degree are accurately matched. The more the pressure loss, the smaller the proportional coefficient, the greater the opening reduction range, the more significant the backflow reduction, and the faster the pressure increase. For example, if the first pneumatic regulating valve final opening is 25.6%, the actual primary exhaust pressure is 0.25 MPa (0.28 MPa lower than the minimum safe pressure), then the adjusted opening = 25.6% x (0.25 ÷ 0.28) ≈ 25.6% x 0.89 ≈ 22.8%, which is 2.8 percentage points lower than the final opening, the cold gas flow backflow is reduced by 2.8%, the primary suction port intake pressure is increased from 0.09 MPa to 0.11 MPa, and the exhaust pressure of the primary compression is increased from 0.25 MPa to 0.29 MPa within 5 seconds, restoring it to the safe range. This effectively solves the problem of insufficient intake air in the secondary head, and the engine exhaust volume is restored from 82% to 98% of the design value.

[0051] The function of the second opening adjusting unit is to cope with the scenario that the primary exhaust pressure is higher than the maximum safe pressure, and to reduce the pressure by adjusting the opening of the second pneumatic regulating valve. The adjustment logic is "second pneumatic regulating valve adjusted opening = second pneumatic regulating valve final opening + (second pneumatic regulating valve final opening - 80%) x (actual primary exhaust pressure - maximum safe pressure) ÷ maximum safe pressure", where "second pneumatic regulating valve final opening" is also a corrected opening value (e.g. 22.6%), "actual primary exhaust pressure" is the value collected by the exhaust pressure sensor (e.g. 0.55 MPa), and "maximum safe pressure" is a value determined by the preset unit (e.g. 0.52 MPa).

[0052] The design logic of the formula needs to be disassembled and explained: "(actual primary exhaust pressure - maximum safe pressure) ÷ maximum safe pressure" is the pressure overrun ratio, which quantifies the degree of pressure exceeding the safety boundary; "(second pneumatic control valve final opening - 80%)" is the opening reference correction term, when the final opening is higher than 80%, the value is positive, which can increase the adjustment range to avoid continuous pressure rise at high opening; when the final opening is lower than 80%, the value is negative, which can reduce the adjustment range to prevent sudden pressure drop caused by excessive reduction of opening; the product of the two and the final opening are added to realize the dynamic adjustment of "pressure overrun degree + current opening state", which ensures that the opening adjustment not only effectively reduces the pressure, but also avoids excessive pressure fluctuation. For example, if the final opening of the second pneumatic control valve is 22.6%, the actual primary exhaust pressure is 0.55 MPa (exceeding the maximum safe pressure by 0.52 MPa), then the pressure overrun ratio = (0.55-0.52) ÷ 0.52 ≈ 0.058, the opening reference correction term = 22.6%-80% = -57.4%, the adjusted opening = 22.6% + (-57.4%) x 0.058 ≈ 22.6%-3.33% ≈ 19.27%, which is 3.33 percentage points lower than the final opening, the secondary exhaust gas supplement amount decreases by 3.33%, the primary exhaust pressure decreases from 0.55 MPa to 0.51 MPa within 6 seconds, and returns to the safety interval, at the same time, the primary head load decreases from 135% of the rated load to 128%, the vibration amplitude decreases from 0.18 mm / s to 0.14 mm / s, which meets the safety operation requirements.

[0053] The function of the third opening degree adjustment unit is to deal with the extreme over-temperature working condition that the actual exhaust temperature exceeds 180 degrees Celsius. At this time, the conventional opening degree adjustment cannot meet the rapid cooling demand, and the extreme valve action needs to be triggered: "close the second pneumatic regulating valve" and "fully open the first pneumatic regulating valve". The "actual exhaust temperature" comes from the exhaust temperature sensor of the first-stage exhaust port. The physical meaning of "closing the second pneumatic regulating valve" is to cut off the air supplement channel from the second-stage exhaust to the first-stage exhaust port, so as to avoid the continuous increase of the air supplement amount, which leads to the increase of the first-stage exhaust pressure and further aggravates the temperature rise. The physical meaning of "fully opening the first pneumatic regulating valve" is to maximize the cold air flow return amount (the return amount is the maximum design value when the opening degree is 100%) from the intermediate cooler outlet to the first-stage suction port, so as to dilute the first-stage suction port inlet temperature by a large amount of cold air, directly reduce the internal energy conversion in the first-stage compression process, and realize rapid cooling. For example, when the actual exhaust temperature reaches 185 degrees Celsius, the third opening degree adjustment unit immediately triggers the action: the second pneumatic regulating valve is quickly closed from 19.27% to 0% (the closing time is ≤1 second), and the air supplement amount is completely cut off; the first pneumatic regulating valve is quickly opened from 22.8% to 100% (the opening time is ≤1 second), the cold air flow return amount is increased from 22.8% of the design value to 100%, the first-stage suction port inlet temperature is reduced from 35 degrees Celsius to 25 degrees Celsius, and the first-stage exhaust temperature is reduced from 185 degrees Celsius to 160 degrees Celsius within 4 seconds and to 140 degrees Celsius within 8 seconds, which effectively avoids the melting and lubrication failure of the sealing element, reduces the production interruption time compared with the traditional "over-temperature shutdown" scheme (the traditional shutdown needs more than 30 minutes to restart, and the scheme does not need to stop to resume), and significantly improves the production continuity.

[0054] Through the synergistic effect of the above-mentioned units, the full-scenario safe control of the first-stage exhaust pressure is realized: in the conventional pressure fluctuation (0.28-0.52 MPa) scenario, the first and second opening degree adjustment units can control the first-stage exhaust pressure fluctuation amplitude within ±0.02 MPa, which is reduced by 60% compared with the traditional scheme (±0.05 MPa); in the pressure overrun scenario (lower than 0.28 MPa or higher than 0.52 MPa), the adjustment response time is shortened to 5-6 seconds, which is improved compared with the traditional scheme (15-20 seconds), and the device damage caused by continuous pressure abnormalities is effectively avoided; in the extreme over-temperature scenario (more than 180 degrees Celsius), the third opening degree adjustment unit can reduce the temperature to the safe range within 8 seconds, and the number of device over-temperature shutdowns is reduced compared with the traditional scheme. Overall, by defining the pressure safety boundary, accurately adjusting in different scenarios, and intervening in extreme working conditions, the pressure safety control capability of the system is improved, which provides a key guarantee for the safe and stable operation of the oil-free screw air compressor under complex pressure working conditions.

[0055] In an embodiment of the present application, the hierarchical adjustment module comprises: a low-load adjustment unit, configured to set the target compression ratio to 0.8 times the design compression ratio and increase the first pneumatic adjustment valve reference opening to a corresponding range (the corresponding range is 60% to 80%) in a low-load mode (the load rate is less than 40%); a medium-load adjustment unit, configured to set the target compression ratio to the design compression ratio and enable proportional-integral-derivative closed-loop control to calculate the opening adjustment amount through the closed-loop control in a medium-load mode (the load rate is between 40% and 70%); by multiplying the proportional coefficient by the inter-stage compression ratio correction amount and adding the integral calculation of the integral coefficient multiplied by the inter-stage compression ratio correction amount; a high-load adjustment unit, configured to set the target compression ratio to 1.2 times the design compression ratio and increase the second pneumatic adjustment valve reference opening to a corresponding range (the corresponding range is 70% to 90%) in a high-load mode (the load rate is greater than 70%); a mode switching unit, configured to use a gradual transition mode to limit the opening change rate to an increase or decrease of no more than 5% per second during mode switching.

[0056] As described above, by constructing a full-load adjustment architecture, i.e., the low-load adjustment unit, the medium-load adjustment unit, and the high-load adjustment unit set the corresponding target compression ratio and valve reference opening range for the load rate less than 40%, 40% to 70%, and greater than 70%, respectively. The medium-load working condition enables proportional-integral-derivative (PID) closed-loop control to optimize the opening adjustment amount. The mode switching unit uses a gradual transition mode to avoid parameter sudden changes, realizes precise adaptation and adjustment of the inter-stage pressure and temperature of the oil-free screw air compressor in the low, medium, and high full-load intervals, solves the problems of insufficient adjustment accuracy and mode switching fluctuations caused by the fixed target compression ratio and single valve reference opening of the basic step adjustment logic in different load working conditions, and ensures that the unit can maintain high compression efficiency and stable inter-stage parameters in the full-load operating range.

[0057] The load rate of the oil-free screw air compressor directly determines the rotor speed of the handpiece, and the speed difference will cause significant changes in the core characteristics of the gas compression process: when the load rate is less than 40%, the rotor speed is lower than 800r / min (taking the rated speed of 1500r / min as an example), and the gas leakage in the rotor meshing gap increases to 15%-20%. The gas cannot fully complete the volume compression during the first-stage compression process. If the designed compression ratio (such as 3.0) is still maintained, the mechanical energy input to the first-stage handpiece will be converted more into gas internal energy, and the first-stage exhaust temperature will easily exceed the safety threshold of 120°C; at the same time, the gas flow velocity in the pipeline decreases under low load, and the base opening of the first pneumatic control valve needs to be increased to increase the cold air flow backflow to effectively suppress the temperature rise. When the load factor is between 40% and 70%, the rotor speed stabilizes at 800-1050r / min, the gas leakage rate drops to 8%-12%, and the first-stage compression efficiency and the second-stage intake demand are in balance. At this time, the inter-stage parameter fluctuations mainly come from small changes in the downstream gas load, and the opening adjustment amount needs to be corrected in real time through PID closed-loop control to maintain the stability of inter-stage pressure and temperature. When the load factor is greater than 70%, the rotor speed exceeds 1050r / min, the first-stage compression volume increases significantly, and the intake volume and intake pressure demand of the second-stage head increase simultaneously. If the design compression ratio is still used, the intake pressure of the second-stage head will be insufficient, and the exhaust pressure of the entire machine will be lower than the design value (for example, the design value of 0.8MPa may drop to 0.7MPa). It is necessary to increase the target compression ratio to increase the first-stage exhaust pressure, and at the same time increase the reference opening of the second pneumatic control valve to optimize the second-stage air supply volume to meet the high compression demand of the second stage. In addition, when switching between different load modes, if the valve opening changes suddenly, it will cause instantaneous fluctuations in the gas flow and pressure in the pipeline, with an amplitude of up to ±0.08MPa or ±10°C, which exceeds the safety tolerance range of the equipment. Therefore, it is necessary to control the opening change rate through a gradual transition method.

[0058] The function of the low load regulation unit is to adapt to the working condition with a load rate less than 40%. The core is to reduce the target compression ratio, and to increase the first pneumatic regulating valve reference opening to reduce the burden of the first stage head and inhibit temperature rise. Among them, the "load rate" comes from the calculation result of the acquisition unit (load rate = current speed ÷ rated speed × 100%, such as current speed 600 r / min, rated speed 1500 r / min, load rate = 600 ÷ 1500 × 100% = 40%, below this value is judged as low load); "design compression ratio" is the rated value (such as 3.0) preset based on the structural parameters of the first stage head rotor tooth number, meshing gap, etc. "Target compression ratio = design compression ratio × 0.8", the value of 0.8 times the design compression ratio is: by reducing the target pressure of the first stage compression, reducing the mechanical energy input in the rotor meshing process, reducing the amount of gas internal energy conversion, thereby inhibiting the temperature rise of the first stage exhaust, for example, when the design compression ratio is 3.0, the target compression ratio = 3.0 × 0.8 = 2.4, the first stage exhaust pressure can be reduced from the design value 0.6 MPa to 0.48 MPa, the load of the first stage head is reduced by 20%, and the temperature rise trend is slowed down. "The first pneumatic regulating valve reference opening is increased to 60% to 80%", the range is determined based on the cold gas flow backflow demand under low load, when the opening is less than 60%, the cold gas flow backflow is insufficient, which cannot effectively dilute the first stage suction port inlet temperature; when it is higher than 80%, the backflow is too large, which causes the first stage suction pressure to be lower than 0.08 MPa, affecting the first stage compression amount, for example, taking the reference opening 70%, which is increased by 20% compared with 50% under the medium load working condition, the cold gas flow backflow is increased by 20%, the first stage suction port inlet temperature is reduced from 35℃ to 30℃, and the first stage exhaust temperature can be controlled within 115℃, avoiding overtemperature.

[0059] The function of the medium load regulation unit is to adapt the load rate to the working condition of 40% to 70%, and the core is to maintain the inter-stage parameter stability through "target compression ratio equal to design compression ratio + PID closed loop control". The physical meaning of "target compression ratio equal to design compression ratio" is that the first stage compression efficiency and the second stage intake demand are balanced under medium load, and maintaining the design compression ratio (such as 3.0) can ensure that the engine displacement reaches the design value (such as 10m³ / min), while avoiding the first stage load being too high or the second stage intake being insufficient. "Enable proportional-integral-derivative closed loop control", the core of this control logic is to dynamically output the opening adjustment amount by real-time calculation of the deviation of the inter-stage compression ratio correction amount, wherein the "proportional coefficient" and "integral coefficient" are preset parameters based on the characteristics of medium load working condition (the proportional coefficient is 0.5 to 0.8, and the integral coefficient is 0.1 to 0.3, such as proportional coefficient 0.6 and integral coefficient 0.2), "opening adjustment amount = proportional coefficient x inter-stage compression ratio correction amount + integral coefficient x ∫ inter-stage compression ratio correction amount dt" (the integral interval is 10 seconds before the current time, which ensures to cover short-term fluctuations). For example, if the inter-stage compression ratio correction amount is 2.0 and the integral value is 5.0, then the opening adjustment amount = 0.6 x 2.0 + 0.2 x 5.0 = 1.2 + 1.0 = 2.2%, that is, the first pneumatic regulating valve is increased by 2.2% based on the reference opening of 50%, and the adjustment is 52.2%. The proportional term quickly responds to the current correction amount, and the integral term eliminates the long-term cumulative deviation, avoiding the deviation superposition caused by traditional open loop control, so that the first stage exhaust temperature fluctuation amplitude is reduced from ±8℃ to ±3℃, and the first stage exhaust pressure fluctuation is reduced from ±0.05MPa to ±0.02MPa, which significantly improves the parameter stability of medium load working condition.

[0060] The function of the high load adjustment unit is to adapt to the working condition with a load rate greater than 70%. The core is to meet the secondary high intake demand by increasing the target compression ratio and the reference opening of the second pneumatic adjusting valve. The target compression ratio = design compression ratio x 1.2. The physical meaning of taking 1.2 times the design compression ratio is that the secondary engine head intake demand increases by 30-40% under high load. By increasing the target compression ratio of the first stage, the exhaust pressure of the first stage can be increased to provide sufficient intake pressure for the secondary engine head. For example, when the design compression ratio is 3.0, the target compression ratio = 3.0 x 1.2 = 3.6, the exhaust pressure of the first stage is increased from 0.6 MPa to 0.72 MPa, and the intake pressure of the secondary engine head meets the design demand (0.6-0.8 MPa). The exhaust pressure of the whole machine is maintained at the design value of 0.8 MPa. The reference opening of the second pneumatic adjusting valve is increased to 70-90%. This range is determined based on the secondary air supply demand under high load. When the opening is less than 70%, the secondary exhaust air supply is insufficient to stabilize the secondary intake pressure. When the opening is greater than 90%, the air supply is too large, causing the exhaust pressure of the first stage to exceed the maximum safe pressure of 0.52 MPa. For example, when the reference opening is 80%, it is increased by 40% compared with the low load condition of 40%, the secondary exhaust air supply is increased by 40%, the exhaust pressure of the first stage is stabilized at 0.72 MPa, and the compression volume of the secondary engine head reaches 110% of the design value, meeting the downstream high gas demand.

[0061] The function of the mode switching unit is to realize smooth transition during mode switching. The core is to avoid parameter sudden change by limiting the opening change rate to an increase or decrease of not more than 5% per second. The trigger condition for mode switching is that the load rate crosses the interval threshold (such as triggering low load to medium load switching when the load rate increases from 38% to 42%, and triggering high load to medium load switching when the load rate decreases from 72% to 68%). The opening change rate is calculated by (target opening - current opening) ÷ transition time. The transition time must satisfy the change rate ≤ 5% / second. For example, when switching from low load to medium load, the current opening of the first pneumatic adjusting valve is 70% and the target opening is 50%. The opening difference is -20%, so the transition time = 20% ÷ 5% / second = 4 seconds, that is, the opening is decreased from 70% to 50% at a rate of 5% per second. When switching from high load to medium load, the current opening of the second pneumatic adjusting valve is 80% and the target opening is 40%. The opening difference is -40%, so the transition time = 40% ÷ 5% / second = 8 seconds, and the opening is adjusted at a rate of 5% per second. By controlling the opening change rate, the gas flow and pressure in the pipeline change smoothly, avoiding pressure shock caused by instantaneous flow fluctuation (such as traditional sudden change with a pressure shock amplitude of 0.1 MPa, and gradual change with a shock amplitude of 0.02 MPa), and preventing temperature overshoot due to flow sudden change (such as traditional sudden change with a temperature overshoot of 5°C, and gradual change with no overshoot). It is ensured that the inter-stage parameters are always within the safe range during mode switching.

[0062] Through the synergy of the above-mentioned units, precise adjustment in the full load range is realized: in the low load range (load rate 30%), the target compression ratio is reduced to 2.4, the first pneumatic regulating valve has a reference opening of 70%, and the first-stage exhaust gas temperature is controlled at 115°C, which is lower than that of the traditional scheme (130°C); in the medium load range (load rate 50%), PID closed-loop control makes the first-stage exhaust gas temperature fluctuate by ±3°C, which is 62.5% lower than that of the traditional open-loop control (±8°C); in the high load range (load rate 80%), the target compression ratio is increased to 3.6, the second pneumatic regulating valve has a reference opening of 80%, and the overall exhaust gas pressure is stabilized at 0.8 MPa; during mode switching, the opening rate is controlled at 5% / second, the pressure fluctuation amplitude is ≤0.02 MPa, and the temperature has no overshoot, which is significantly optimized compared with the traditional sudden change scheme (pressure fluctuation 0.1 MPa, temperature overshoot 5°C). Overall, through load partitioning adaptation, closed-loop fine adjustment, and smooth switching, the system's adjustment capability in the full load range is improved, providing key support for efficient and stable operation of the unit in all operating conditions.

[0063] In an embodiment of the present application, the temperature and pressure monitoring module further comprises: A third calculation unit for calculating the change acceleration of the first-stage exhaust gas temperature, the change acceleration being calculated by the change value of the temperature change rate per unit time, with the unit being degrees Celsius per second squared; A generation unit for generating a lead adjustment signal when the current change acceleration is greater than the preset maximum change acceleration (the preset maximum change acceleration being 0.5 degrees Celsius per second squared); the signal value being calculated by multiplying the acceleration by the current opening and then by a coefficient of 0.05 to 0.15; A superposition unit for superimposing the lead adjustment signal on the first pneumatic regulating valve opening command; the new opening being equal to the current opening plus the lead adjustment signal; A protection unit for starting the anti-overadjustment protection and freezing the first pneumatic regulating valve opening for 10 to 15 seconds when the current change acceleration is less than the preset minimum change acceleration (the preset minimum change acceleration being -0.3 degrees Celsius per second squared).

[0064] As described above, by constructing a predictive temperature control logic, i.e., the third calculation unit calculates the temperature change acceleration based on the difference of the temperature change rate per unit time, the generation unit generates a lead adjustment signal when the acceleration exceeds the preset maximum value, the superposition unit integrates the signal into the first pneumatic regulating valve opening command to intervene in the temperature change trend in advance, and the protection unit freezes the opening to avoid overadjustment when the acceleration is lower than the preset minimum value, the speed of the first-stage exhaust gas temperature change trend is accurately captured and predictively controlled, the adjustment lag and overadjustment risk problems existing in the temperature change rate compensation-based scheme are solved, the responsiveness and adjustment stability of the system to temperature dynamic changes are further improved, and the inter-stage temperature is ensured to be always stable in the safe and efficient range.

[0065] The core function of the third calculation unit is to calculate the acceleration of the change of the primary exhaust temperature in real time. The parameter is calculated by "acceleration of change = (current temperature change rate - previous temperature change rate) ÷ time interval". The "current temperature change rate" and the "previous temperature change rate" are both from the monitoring results of the detection unit, and the "time interval" is the time difference between the two change rates (fixed at 1 second, which can accurately reflect the dynamic change of the change rate and avoid data noise interference due to too short interval). The unit of the calculation result is degree Celsius per second square. The acceleration of change directly quantifies the speed of the temperature change trend. For example, if the current temperature change rate is 2.6°C / s and the previous change rate (1 second ago) is 2.0°C / s, then the acceleration of change = (2.6-2.0) ÷ 1 = 0.6°C / s², indicating that the temperature rising speed is accelerating. If the current change rate is 1.4°C / s and the previous change rate is 2.0°C / s, then the acceleration of change = (1.4-2.0) ÷ 1 = -0.6°C / s², indicating that the temperature rising speed is rapidly slowing down. The acceleration of change data calculated by this unit is transmitted to the generation unit and the protection unit in real time. Its significance is to upgrade the dynamic trend of temperature change from the first-order change rate to the second-order acceleration, providing a core basis for the subsequent advanced regulation and over-regulation protection, and avoiding the traditional lagging judgment relying only on the change rate.

[0066] The function of the generation unit is to generate an advanced regulation signal when the acceleration of change exceeds the preset maximum value, to enhance the opening degree adjustment in advance, and to offset the regulation lag caused by the intensification of temperature change trend. The preset maximum value of the acceleration of change is set to 0.5°C / s², which is determined based on the temperature regulation inertia characteristics of the oil-free screw air compressor. When the acceleration exceeds 0.5°C / s², if there is no advance intervention, the temperature change rate will rise to more than 2.5°C / s within 5 seconds, and the primary exhaust temperature will exceed the safety threshold. The "advanced regulation signal value = acceleration of change × current opening degree × preset coefficient", where "current opening degree" is the corrected first pneumatic regulating valve opening degree (such as 25%, which has been compensated for load rate and thermal inertia), and "preset coefficient" is valued at 0.05 to 0.15 (such as 0.1, which ensures that the signal value can effectively intervene in advance and avoid excessive regulation). The greater the acceleration of change, the more obvious the intensification of temperature change trend, and the greater the advanced signal required. The greater the current opening degree, the stronger the current cooling capacity, and the corresponding signal amplitude is required to maintain stable regulation. For example, if the acceleration of change is 0.6°C / s² (exceeding the preset maximum value of 0.5°C / s²), the current opening degree is 25%, and the coefficient is 0.1, then the advanced regulation signal value = 0.6 × 25% × 0.1 = 1.5%, which indicates that the current opening degree needs to be increased by 1.5% to enhance the cold air flow volume in advance and cope with the intensifying temperature rising trend.

[0067] The function of the superposition unit is to integrate the advance adjustment signal into the opening command of the first pneumatic regulating valve to form the final opening value for execution. The adjustment logic is "first pneumatic regulating valve new opening = current opening + advance adjustment signal value". Wherein "current opening" is the corrected opening (such as 25%), "advance adjustment signal value" is the calculation result of the generation unit (such as 1.5%), then new opening = 25% + 1.5% = 26.5%. By directly superimposing the advance signal on the current opening, the valve is advanced to increase the opening (or decrease the opening in the downward trend), so that the cold gas flow rate adjustment precedes the temperature change trend, thereby offsetting the lag caused by the adjustment inertia. For example, without superimposing the signal, the current opening of 25% can only cope with a change rate of 2.0℃ / s, and the temperature will rise to 130℃ after 5 seconds; after superimposing a signal of 1.5%, the opening of 26.5% can cope with a change rate of 2.6℃ / s, and the temperature will only rise to 125℃ after 5 seconds, avoiding overshoot. This superposition mechanism can reduce the temperature overshoot from the traditional 8-12℃ to 3-4℃, shorten the adjustment lag time, and significantly improve the timeliness of the response.

[0068] The function of the protection unit is to deal with the risk of over-regulation when the change acceleration is lower than the preset minimum value. When the change acceleration is less than -0.3℃ / s² (the preset minimum value is determined based on the compression efficiency requirement, when the acceleration is lower than -0.3℃ / s², the temperature drop speed will increase from 1℃ / s to 1.9℃ / s within 3 seconds, which is easy to cause the temperature to drop below the high-efficiency compression interval of 100℃), the anti-over-regulation protection is started, and the first pneumatic regulating valve opening is frozen for 10 to 15 seconds (such as 12 seconds, which is determined by test, which can both suppress the temperature from falling too much and avoid freezing for too long to cause new fluctuations). For example, the change acceleration is -0.4℃ / s² (lower than -0.3℃ / s²), and the current first pneumatic regulating valve opening is 26.5%, then the protection unit immediately freezes the opening for 12 seconds, and the opening is not adjusted according to the temperature change during this period. When the temperature drop trend is too fast, the adjustment logic of continuously increasing the flow rate (or reducing the opening) is cut off by freezing the opening, so as to avoid the temperature from falling too much. If it is not frozen, the opening will continue to increase to 28%, the flow rate will further increase, and the temperature will drop from 120℃ to 95℃ within 12 seconds, and the primary compression efficiency will decrease from 85% to 70%; after freezing, the temperature only drops to 105℃, still maintaining in the high-efficiency interval, and the compression efficiency remains in the high-efficiency interval, while avoiding the parameter fluctuations caused by the reverse regulation due to the temperature being too low.

[0069] Through the synergy of the above-mentioned units, the trend prediction control of the primary exhaust temperature change is realized: in the temperature rising trend intensification scene (acceleration 0.6℃ / s²), the temperature overshoot amplitude is reduced by the advance adjustment signal, and the adjustment lag time is shortened; in the temperature drop trend too fast scene (acceleration-0.4℃ / s²), the temperature minimum value is maintained at 105℃ by the anti-overadjustment protection, and the compression efficiency loss is reduced; in the conventional fluctuation scene (acceleration-0.2 to 0.4℃ / s²), the system does not need to start the advance or protection mechanism, only maintains the original compensation logic, and ensures the adjustment stability. Overall, through the monitoring and predictive control of the temperature change acceleration, the technical gap of the traditional adjustment based on the change rate is filled, the control accuracy and stability of the system in the dynamic temperature change scene are further improved, and more precise temperature protection is provided for the efficient and safe operation of the unit.

[0070] In an embodiment of the present application, a calculation module is further included, which comprises: A fourth calculation unit is configured to obtain the primary suction temperature and the primary suction pressure through the suction temperature sensor and the suction pressure sensor, obtain the primary exhaust temperature and the primary exhaust pressure through the exhaust temperature sensor and the exhaust pressure sensor, and calculate the actual compression ratio based on the primary suction temperature, the primary exhaust temperature, the primary suction pressure and the primary exhaust pressure; The specific calculation process is: the actual compression ratio is calculated based on the ideal gas equation, the primary suction temperature, the primary exhaust temperature, the primary suction pressure and the primary exhaust pressure are first obtained, and then the exhaust pressure is divided by the suction pressure, and then multiplied by the (adiabatic index divided by 1 minus adiabatic index) power of (suction temperature divided by exhaust temperature), and the adiabatic index is valued between 1.30 and 1.40; A fifth calculation unit is configured to calculate the compression efficiency based on the actual compression ratio: The compression efficiency is calculated by (the (1 minus adiabatic index divided by adiabatic index) power of the actual compression ratio minus 1) divided by (the (1 minus adiabatic index divided by adiabatic index) power of the design compression ratio minus 1) and multiplied by 100%; A sixth calculation unit is configured to trigger the efficiency optimization mode when the compression efficiency is less than a preset efficiency value (the preset efficiency value is 75%), and the specific optimization method is to reduce the set temperature value by 10 to 20 degrees Celsius and reduce the design compression ratio by 0.2 to 0.5.

[0071] As described above, by constructing a closed-loop efficiency control system, i.e., the fourth calculation unit calculates the actual compression ratio based on the temperature and pressure parameters of the primary suction and exhaust, the fifth calculation unit quantifies the compression efficiency by comparing the actual compression ratio with the design compression ratio, and the sixth calculation unit triggers the efficiency optimization mode when the compression efficiency is lower than the preset value, adjusts the set temperature value and the design compression ratio, realizes real-time monitoring and dynamic optimization of the compression efficiency of the oil-free screw air compressor, solves the problem that the above content only focuses on the control of inter-stage pressure and temperature, and does not actively control the compression efficiency, ensures that the unit always runs in the high-efficiency interval while maintaining stable inter-stage parameters, reduces invalid energy consumption, and improves the operation economy.

[0072] The core function of the fourth calculation unit is to calculate the actual compression ratio based on the temperature and pressure parameters of the primary suction and exhaust. This calculation process is based on the ideal gas state equation as the theoretical basis, and fully considers the temperature and pressure coupling relationship in the gas compression process. Among them, the "primary suction temperature" and "primary suction pressure" required for calculation come from the suction temperature sensor (collection accuracy ±0.5℃) and suction pressure sensor (collection accuracy ±0.01MPa) installed at the primary suction port in the temperature and pressure monitoring module, and the "primary exhaust temperature" and "primary exhaust pressure" come from the exhaust temperature sensor and exhaust pressure sensor installed at the primary exhaust port. The collection frequency of all parameters is 1 time / second, which ensures that the data reflects the gas compression state in real time. The specific calculation formula is: Among them, represents the actual compression ratio, represents the primary exhaust pressure, represents the primary suction pressure, represents the primary suction temperature, represents the primary exhaust temperature, represents the adiabatic index; ​wherein the "adiabatic index" takes a value of 1.30 to 1.40 (for air medium, the value is determined by experiment, 1.35 is a typical value, reflecting the heat transfer characteristics of air during compression, the larger the adiabatic index, the more significant the temperature rise during gas compression). During the body compression process, there is a coupling relationship between pressure change and temperature change, and only by pressure ratio cannot accurately reflect the actual compression degree, and the temperature ratio and the adiabatic index correction need to be introduced to obtain the compression ratio that fits the actual compression process. For example, the first-stage suction temperature is 300K (27℃), the first-stage suction pressure is 0.1MPa, the first-stage exhaust temperature is 360K (87℃), the first-stage exhaust pressure is 0.6MPa, and the adiabatic index is 1.35, then the actual compression ratio = (0.6 ÷ 0.1) × (300 ÷ 360)^(1.35 ÷ (1.35-1)) = 6 × (0.833)^3.857 ≈ 6 × 0.42 ≈ 2.52, which indicates that the current actual compression ratio is lower than the design value of 3.0, and there is a risk of efficiency decline.

[0073] The function of the fifth calculation unit is to quantify the current compression efficiency by comparing the actual compression ratio with the design compression ratio, and to establish a direct correlation between efficiency and compression ratio. The calculation logic is "compression efficiency = ((actual compression ratio^((adiabatic index-1) ÷ adiabatic index))-1) ÷ ((design compression ratio^((adiabatic index-1) ÷ adiabatic index))-1) × 100%", wherein the "design compression ratio" is the rated compression ratio determined in the design stage of the unit (such as 3.0, stored in the PLC control module parameter library), and the "adiabatic index" is the same as that of the fourth calculation unit (such as 1.35). The compression efficiency is essentially the ratio of the useful work consumed in the actual compression process to the useful work consumed in the ideal compression process (under the design compression ratio), and by substituting the actual and design compression ratios into the adiabatic compression work formula, the efficiency difference can be directly quantified. The closer the actual compression ratio is to the design value, the closer the efficiency is to 100%; the greater the deviation of the actual compression ratio, the lower the efficiency. For example, the actual compression ratio is 2.52, the design compression ratio is 3.0, and the adiabatic index is 1.35, then the actual compression ratio term = (2.52^((1.35-1) ÷ 1.35))-1 = (2.52^0.259)-1 ≈ 1.26-1 = 0.26, the design compression ratio term = (3.0^0.259)-1 ≈ 1.32-1 = 0.32, and the compression efficiency = (0.26 ÷ 0.32) × 100% ≈ 81.25%, which indicates that the current efficiency is at a relatively high level, but there is still room for optimization; if the actual compression ratio drops to 2.2, the actual compression ratio term = (2.2^0.259)-1 ≈ 1.21-1 = 0.21, and the compression efficiency = (0.21 ÷ 0.32) × 100% ≈ 65.62%, which is lower than the preset efficiency value of 75%, and optimization needs to be triggered.

[0074] The function of the sixth calculation unit is to trigger the efficiency optimization mode when the compression efficiency is less than the preset efficiency value of 75%, and to restore the efficiency by adjusting the core control parameters. The preset efficiency value of 75% is determined based on the energy efficiency and equipment safety boundary of the oil-free screw air compressor. When the efficiency is lower than 75%, the energy consumption per unit of exhaust volume will exceed the industry energy efficiency level II standard (GB19153-2019), and long-term operation will exacerbate the wear of the machine head. The optimization method is to reduce the set temperature value by 10 to 20 degrees Celsius and to reduce the design compression ratio by 0.2 to 0.5. The two work together: reducing the set temperature value (such as from 120°C to 105°C) can enhance the cold gas flow backflow adjustment space of the first pneumatic regulating valve, inhibit the temperature rise caused by the decrease in efficiency, and avoid further deterioration of the leakage caused by the temperature rise. Reducing the design compression ratio (such as from 3.0 to 2.7) can reduce the compression load of the primary machine head, so that the actual compression ratio is closer to the adjusted design compression ratio, reducing the energy waste caused by the deviation of the compression ratio. For example, when the compression efficiency decreases to 65.62%, the optimization mode is triggered: the set temperature value is reduced from 120°C to 105°C, and the design compression ratio is reduced from 3.0 to 2.7. At this time, the fourth calculation unit recalculates the actual compression ratio to 2.5 (because the design compression ratio is reduced, the deviation between the actual and the design is reduced), and the fifth calculation unit calculates the compression efficiency = ((2.5^0.259)-1) / ((2.7^0.259)-1)*100%≈(1.24-1) / (1.27-1)*100%≈0.24 / 0.27*100%≈88.89%, the efficiency is significantly improved to the high efficiency interval, the energy consumption per unit of exhaust volume is reduced by 16%, and the primary exhaust temperature is reduced from 135°C to 110°C, avoiding the rotor from being stuck due to high temperature.

[0075] Through the synergistic effect of the above-mentioned units, the compression efficiency is controlled throughout the cycle: in the efficiency monitoring stage, the fourth and fifth calculation units can quantitatively monitor the actual compression ratio and efficiency in real time, with a monitoring accuracy of ±2%, which realizes visual management of efficiency compared with traditional monitoring schemes; in the efficiency optimization stage, the parameter adjustment of the sixth calculation unit can improve the efficiency and reduce the energy consumption without stopping the machine, avoiding production interruption caused by maintenance; in the long-term operation stage, the module can dynamically adapt to the decrease in actual compression ratio caused by machine head wear, prolonging the high-efficiency operation period of the unit (from the traditional 2000 hours to 3500 hours) and reducing the maintenance frequency and cost. Overall, through efficiency quantification and dynamic optimization, the blank in efficiency control of traditional control schemes is filled, realizing the dual goals of stable inter-stage parameters and high-efficiency compression, and providing key technical support for the economic operation of the unit.

[0076] In one embodiment of the present application, the present application also discloses a computer device comprising a memory and a processor, wherein the memory stores a computer program, and the processor executes the computer program to realize the operation of the inter-stage pressure and temperature flexible control system of the oil-free screw air compressor.

[0077] In one embodiment of the present application, the present application also discloses a computer readable storage medium, which stores a computer program, and the computer program is executed by a processor to realize the operation of the inter-stage pressure and temperature flexible control system of the oil-free screw air compressor.

[0078] It can be understood by those skilled in the art that all or part of the processes of the above-mentioned embodiments can be completed by a computer program instructing related hardware, and the computer program can be stored in a non-volatile computer readable storage medium. When the computer program is executed, it can include the processes of the above-mentioned embodiments. Any reference to memory, storage, value library or other medium provided by the present application and used in the embodiments can include non-volatile and / or volatile memory. Non-volatile memory can include read-only memory (ROM), programmable ROM (PROM), electrically programmable ROM (EPROM), electrically erasable programmable ROM (EEPROM) or flash memory. Volatile memory can include random access memory (RAM) or external cache memory. As an illustration but not limitation, RAM is available in various forms, such as static RAM (SRAM), dynamic RAM (DRAM), synchronous DRAM (SDRAM), double data rate SDRAM (SSRSDRAM), enhanced SDRAM (ESDRAM), synchronous link (Synchlink) DRAM (SLDRAM), memory bus (Rambus) direct RAM (RDRAM), direct memory bus dynamic RAM (DRDRAM), and memory bus dynamic RAM (RDRAM) and the like.

[0079] It should be noted that in this document, the terms "comprising", "including" or any other variant thereof are intended to cover non-exclusive inclusions, so that a process, device, article or method including a series of elements not only includes those elements, but also includes other elements not explicitly listed, or includes elements inherent to such a process, device, article or method. Without more limitations, the element defined by the statement "including a" does not exclude the presence of another identical element in the process, device, article or method including the element.

[0080] The above merely describes the preferred embodiments of the present application, and is not intended to limit the scope of the present application, and any equivalent results or equivalent process transformations made by using the content of the present application specification and drawings, or directly or indirectly applied in other related technical fields, are also included in the protection scope of the present application.

Claims

1. A flexible control system for interstage pressure and temperature of an oil-free screw air compressor, characterized in that: include: A first pneumatic regulating module is used to add a first flexible pipeline between the intercooler outlet and the first-stage air intake, wherein a first pneumatic regulating valve is installed in the middle of the first flexible pipeline; A second pneumatic regulating module is used to add a second flexible pipeline between the secondary exhaust port and the aftercooler inlet, and a second pneumatic regulating valve is installed in the middle of the second flexible pipeline; Temperature and pressure monitoring module, used to install the intake temperature sensor and intake pressure sensor at the first-stage intake port, the exhaust temperature sensor and exhaust pressure sensor at the first-stage exhaust port, and the total exhaust temperature sensor at the exhaust port of the whole machine; A PLC control module is used to receive sensor data from the temperature and pressure monitoring module in real time, and calculate the inter-stage compression ratio correction value based on the difference between the actual exhaust temperature and the set temperature value and the deviation between the actual pressure and the design compression ratio; A staged regulation module is used to dynamically regulate the openings of the first pneumatic regulating valve and the second pneumatic regulating valve based on the inter-stage compression ratio correction under different load pressures; The opening adjustment module is used to generate a valve opening instruction according to the inter-stage compression ratio correction amount. The opening of the first pneumatic control valve is calculated based on the negative exponential function of the correction amount, and the opening of the second pneumatic control valve is calculated based on the positive exponential function of the correction amount.

2. The interstage pressure and temperature flexible control system of the oil-free screw air compressor according to claim 1, characterized in that: The PLC control module includes: The collecting unit is used to collect the load rate of the air compressor in real time, where the load rate of the air compressor is obtained by calculating the current speed and the rated speed; A first calculation unit is used to calculate a load correction coefficient, wherein the load correction coefficient is determined based on the load rate in sections; A first correction unit is used to correct the opening of the first pneumatic control valve by multiplying the initial opening by a load correction coefficient to obtain the corrected opening of the first pneumatic control valve; a second correction unit, configured to correct the opening of the second pneumatic control valve by multiplying the initial opening by 2 and subtracting the load correction coefficient to obtain a corrected opening of the second pneumatic control valve; The forced correction unit is used to force the corrected opening of the first pneumatic control valve to increase by 20% to 35% when the actual exhaust temperature exceeds the alarm threshold.

3. The interstage pressure and temperature flexible control system of the oil-free screw air compressor according to claim 1, characterized in that: The temperature and pressure monitoring module includes: A detection unit, configured to monitor a rate of change of the first-stage exhaust gas temperature, wherein the rate of change of the first-stage exhaust gas temperature is calculated based on a temperature change value per unit time; a second calculation unit, configured to calculate a thermal inertia compensation coefficient, wherein the calculated thermal inertia compensation coefficient is determined based on a temperature change rate; an opening adjustment unit, configured to apply a thermal inertia compensation coefficient to valve opening adjustment, wherein the opening of the first pneumatic control valve is calculated by multiplying the corrected opening of the first pneumatic control valve by the compensation coefficient, and the opening of the second pneumatic control valve is calculated by multiplying the corrected opening of the second pneumatic control valve by 2 and subtracting the thermal inertia compensation coefficient; The emergency cooling unit is used to trigger the emergency cooling mode when the temperature change rate is greater than a preset value.

4. The interstage pressure and temperature flexible control system of the oil-free screw air compressor according to claim 1, characterized in that: The PLC control module also includes: The preset unit is used to preset the first-level exhaust pressure safety range, where the minimum safety pressure is 0.7 times the design pressure and the maximum safety pressure is 1.3 times the design pressure; A first opening adjustment unit is used to adjust the opening of the first pneumatic control valve when the actual first-stage exhaust pressure is lower than the minimum safety pressure; The second opening adjustment unit is used to adjust the opening of the second pneumatic control valve when the actual first-stage exhaust pressure is greater than the maximum safety pressure: The third opening adjustment unit is used to close the second pneumatic control valve and fully open the first pneumatic control valve when the actual exhaust temperature exceeds 180 degrees Celsius.

5. The interstage pressure and temperature flexible control system of the oil-free screw air compressor according to claim 1, characterized in that: The hierarchical adjustment module includes: A low-load adjustment unit is used to set the target compression ratio to 0.8 times the design compression ratio in the low-load mode and increase the reference opening of the first pneumatic control valve to a corresponding range; a medium load adjustment unit, configured to set a target compression ratio equal to a design compression ratio and enable closed-loop control in a medium load mode, wherein an opening adjustment amount is calculated by the closed-loop control; Calculated by multiplying the proportional coefficient by the inter-stage compression ratio correction amount plus the integral coefficient by the inter-stage compression ratio correction amount; A high-load adjustment unit is used to set the target compression ratio to 1.2 times the design compression ratio in high-load mode and increase the reference opening of the second pneumatic control valve to a corresponding range; The mode switching unit is used to adopt a gradual transition method when switching modes, and the opening change rate is limited to an increase or decrease of no more than 5% per second.

6. The interstage pressure and temperature flexible control system of the oil-free screw air compressor according to claim 1, characterized in that: The temperature and pressure monitoring module also includes: a third calculation unit, configured to calculate a change acceleration of the first-stage exhaust gas temperature, wherein the change acceleration is calculated by a change value of the temperature change rate per unit time; A generating unit, configured to generate an advance adjustment signal when the current change acceleration is greater than a preset maximum value of the change acceleration; a superposition unit, configured to superimpose the advance adjustment signal onto the opening instruction of the first pneumatic control valve; The protection unit is used to start the anti-overmodulation protection when the current change acceleration is less than the preset minimum value of the change acceleration.

7. The interstage pressure and temperature flexible control system of the oil-free screw air compressor according to claim 1, characterized in that: Also included is a calculation module, the calculation module including: a fourth calculation unit, configured to obtain a first-stage intake temperature and a first-stage intake pressure via an intake temperature sensor and an intake pressure sensor, obtain a first-stage exhaust temperature and a first-stage exhaust pressure via an exhaust temperature sensor and an exhaust pressure sensor, and calculate an actual compression ratio based on the first-stage intake temperature, the first-stage exhaust temperature, the first-stage intake pressure, and the first-stage exhaust pressure; The fifth calculation unit is used to calculate the compression efficiency according to the actual compression ratio: The sixth calculation unit is configured to trigger an efficiency optimization mode when the compression efficiency is less than a preset efficiency value.

8. A computer device comprising a memory and a processor, wherein the memory stores a computer program, wherein: When the processor executes the computer program, the system according to any one of claims 1 to 7 is implemented.

9. A computer-readable storage medium having a computer program stored thereon, characterized in that: When the computer program is executed by a processor, the system according to any one of claims 1 to 7 is implemented.

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