Inter-stage pressure temperature flexible control system of oil-free screw air compressor

Abstract

The present application relates to air compressor regulation and control technology field, specifically disclose a kind of interstage pressure temperature flexible control system of oil-free screw air compressor, for the fixed interstage parameter of existing system, thermal inertia response lag, compression efficiency is not real-time management and control, poor load adaptability, easy to cause over-temperature safety hazard, energy-saving potential is insufficient, operating economy is low The problem, the system includes first pneumatic regulating module, second pneumatic regulating module, temperature and pressure monitoring module, PLC control module, hierarchical regulating module and opening degree regulating module, by constructing flexible pipeline regulating loop, multiple measuring point monitoring, dynamic correction amount calculation and hierarchical opening degree control The collaborative control architecture of regulation, realize the accurate flexible control to first and second interstage pressure and temperature, further solve the problem that the minimum energy-saving speed is limited due to the too high first-stage exhaust temperature of oil-free screw air compressor under low-speed load, achieve the goal of safe operation of unit at lower speed, further improve energy-saving effect.

Description

Technical Field

[0001] This invention relates to the field of air compressor control technology, and in particular to a flexible interstage pressure and temperature control system for an oil-free screw air compressor. Background Technology

[0002] Oil-free screw air compressors are widely used in fields with high requirements for compressed air quality, such as chemical, electronic, and medical industries, due to their characteristics of no oil pollution and stable operation. The interstage pressure and temperature parameters between the first and second stage compressor heads directly determine the compression efficiency, equipment safety, and energy-saving effect, and are the core control objects.

[0003] There are some shortcomings in the interstage parameter control of existing oil-free screw air compressors:

[0004] First, the use of fixed pipelines connecting the first and second stage compressor heads and the cooler, with the interstage compression ratio fixed at the design stage, cannot adapt to load fluctuations. Under low load, the rotor speed is low, gas leakage increases, and the first stage exhaust temperature easily exceeds the safety threshold, limiting the unit's minimum energy-saving speed. Under high load, the second stage intake pressure is insufficient, the actual compression ratio of the second stage increases, and the second stage exhaust temperature easily exceeds the safety threshold. Second, relying on the intercooler for passive cooling, lacking an active adjustment loop, coupled with the thermal inertia of the gas compression process, results in a delayed response to rapid temperature changes, easily leading to temperature overshoot. Third, the control logic is not linked to compression efficiency, relying only on empirically set temperature and pressure thresholds without real-time calculation of the actual compression ratio and efficiency. When the efficiency is below 75%, the original parameters are maintained, increasing energy consumption per unit exhaust volume by more than 13%.

[0005] Therefore, there is an urgent need for a flexible interstage pressure and temperature control system for oil-free screw air compressors to solve the above problems. Summary of the Invention

[0006] The purpose of this invention is to provide a flexible interstage pressure and temperature control system for an oil-free screw air compressor, comprising:

[0007] The first pneumatic adjustment module is used to add a first flexible pipeline between the outlet of the intercooler and the first-stage intake port, and a first pneumatic adjustment valve is installed in the middle of the first flexible pipeline.

[0008] The second pneumatic adjustment module is used to add a second flexible pipeline between the secondary exhaust port and the aftercooler inlet, and a second pneumatic adjustment valve is installed in the middle of the second flexible pipeline;

[0009] The temperature and pressure monitoring module is used to install an intake temperature sensor and an intake pressure sensor at the first-stage intake port, an exhaust temperature sensor and an exhaust pressure sensor at the first-stage exhaust port, and a total exhaust temperature sensor at the exhaust port of the whole machine.

[0010] The PLC control module is used to receive sensor data from the temperature and pressure monitoring module in real time, and to calculate the interstage compression ratio correction based on the difference between the actual exhaust temperature and the set temperature value, as well as the deviation between the actual pressure and the designed compression ratio.

[0011] The graded adjustment module is used to dynamically adjust the opening of the first pneumatic control valve and the second pneumatic control valve based on the interstage compression ratio correction under different load pressures.

[0012] The valve opening adjustment module is used to generate valve opening commands based on the interstage compression ratio correction. The opening of the first pneumatic control valve is calculated based on a negative exponential function of the correction, and the opening of the second pneumatic control valve is calculated based on a positive exponential function of the correction.

[0013] Furthermore, the PLC control module includes:

[0014] The data acquisition unit is used to acquire the load rate of the air compressor in real time. The load rate of the air compressor is calculated by comparing the current speed with the rated speed.

[0015] The first calculation unit is used to calculate the load correction factor, which is determined based on the load rate segmentation.

[0016] The first correction unit is used to correct the opening degree of the first pneumatic control valve by multiplying the initial opening degree by the load correction coefficient to obtain the corrected opening degree of the first pneumatic control valve.

[0017] The second correction unit is used to correct the opening degree of the second pneumatic control valve. The corrected opening degree of the second pneumatic control valve is obtained by multiplying the initial opening degree by 2 and subtracting the load correction coefficient.

[0018] The forced correction unit is used to forcibly increase the opening of the corrected first pneumatic regulating valve by 20% to 35% when the actual exhaust temperature exceeds the alarm threshold.

[0019] Furthermore, the temperature and pressure monitoring module includes:

[0020] The detection unit is used to monitor the rate of change of the primary exhaust temperature, which is calculated based on the temperature change value per unit time.

[0021] The second calculation unit is used to calculate the thermal inertia compensation coefficient, which is determined based on the rate of temperature change.

[0022] The valve opening adjustment unit is used to apply the thermal inertia compensation coefficient to the valve opening adjustment. The opening of the first pneumatic regulating valve is calculated by multiplying the corrected opening of the first pneumatic regulating valve by the compensation coefficient. The opening of the second pneumatic regulating valve is calculated by multiplying the corrected opening of the second pneumatic regulating valve by 2 and subtracting the difference from the thermal inertia compensation coefficient.

[0023] The emergency cooling unit is used to trigger the emergency cooling mode when the rate of temperature change exceeds a preset value.

[0024] Furthermore, the PLC control module also includes:

[0025] The preset unit is used to preset the first-level 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;

[0026] The first opening adjustment unit is used to adjust the opening of the first pneumatic regulating valve when the actual first-stage exhaust pressure is less than the minimum safe pressure.

[0027] The second opening adjustment unit is used to adjust the opening of the second pneumatic regulating valve when the actual first-stage exhaust pressure is greater than the maximum safe pressure.

[0028] The third opening adjustment unit is used to close the second pneumatic regulating valve and fully open the first pneumatic regulating valve when the actual exhaust temperature exceeds 180 degrees Celsius.

[0029] Furthermore, the graded adjustment module includes:

[0030] The low-load regulating unit is used to set the target compression ratio to 0.8 times the design compression ratio in low-load mode and increase the reference opening of the first pneumatic regulating valve to the corresponding range.

[0031] The medium load adjustment unit is used to set the target compression ratio to equal the design compression ratio in medium load mode and enable closed-loop control. The opening adjustment amount is calculated through the closed-loop control.

[0032] It is calculated by multiplying the proportional coefficient by the interstage compression ratio correction and then integrating the integral coefficient by the interstage compression ratio correction.

[0033] The high-load regulating 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 regulating valve to the corresponding range.

[0034] The mode switching unit is used to employ a gradual transition method during mode switching, with the opening change rate limited to an increase or decrease of no more than 5% per second.

[0035] Furthermore, the temperature and pressure monitoring module also includes:

[0036] The third calculation unit is used to calculate the acceleration of the change in the first-stage exhaust temperature, which is calculated by the change in the rate of temperature change per unit time.

[0037] The generation unit is used to generate an advance adjustment signal when the current change acceleration is greater than the preset maximum value of the change acceleration;

[0038] The superposition unit is used to superimpose the advance adjustment signal onto the opening command of the first pneumatic control valve;

[0039] The protection unit is used to activate over-adjustment protection when the current changing acceleration is less than the preset minimum value of the changing acceleration.

[0040] Furthermore, it also includes a computing module, which comprises:

[0041] The fourth calculation unit is used to obtain the first-stage intake temperature and first-stage intake pressure through the intake temperature sensor and the intake pressure sensor, and to obtain the first-stage exhaust temperature and first-stage exhaust pressure through the exhaust temperature sensor and the exhaust pressure sensor, and to calculate the actual compression ratio based on the first-stage intake temperature, first-stage exhaust temperature, first-stage intake pressure and first-stage exhaust pressure.

[0042] The fifth calculation unit is used to calculate the compression efficiency based on the actual compression ratio.

[0043] The sixth calculation unit is used to trigger the efficiency optimization mode when the compression efficiency is less than the preset efficiency value.

[0044] Furthermore, the present invention also discloses a computer device, including a memory and a processor, wherein the memory stores a computer program, and the processor executes the computer program to realize the operation of an interstage pressure and temperature flexible control system for an oil-free screw air compressor.

[0045] Furthermore, the present invention also discloses a computer-readable storage medium storing a computer program thereon, which, when executed by a processor, enables the operation of an interstage pressure and temperature flexible control system for an oil-free screw air compressor.

[0046] The beneficial effects of this application are as follows:

[0047] This invention employs a collaborative control architecture that integrates flexible pipeline regulation loops, multi-point monitoring, dynamic correction calculation, and tiered opening control. Specifically, a first and second pneumatic regulation module constructs flexible pipelines between the intercooler outlet and the first-stage intake port, and the second-stage exhaust port and aftercooler inlet, respectively. A temperature and pressure monitoring module collects temperature and pressure parameters from the first-stage intake, first-stage exhaust, and overall exhaust. A PLC control module calculates the interstage compression ratio correction based on temperature and pressure differences. A tiered regulation module adapts to different load and pressure conditions. Finally, an opening control module regulates the opening of two pneumatic regulating valves according to an exponential function, achieving precise and flexible control of pressure and temperature between the first and second stages. This solves the problem of excessively high first-stage exhaust temperature limiting the minimum energy-saving speed in oil-free screw air compressors under low-speed loads, enabling safe operation of the unit at lower speeds and further improving energy efficiency. Attached Figure Description

[0048] Figure 1 This is a schematic diagram of the system structure proposed in an embodiment of the present invention.

[0049] The realization of the purpose, functional features and advantages of this application will be further explained in conjunction with the embodiments and with reference to the accompanying drawings. Detailed Implementation

[0050] It should be understood that the specific embodiments described herein are merely illustrative of the invention and are not intended to limit the invention.

[0051] like Figure 1 As shown, this application provides a flexible interstage pressure and temperature control system for an oil-free screw air compressor, comprising:

[0052] The first pneumatic regulating module 1 is used to add a first flexible pipeline between the outlet of the intercooler and the first-stage intake port. A first pneumatic regulating valve is installed in the middle of the first flexible pipeline. The valve diameter is determined based on the gas flow rate under the lowest load condition and the maximum allowable flow velocity of the pipeline. Specifically, the diameter value is calculated by multiplying the flow correction coefficient by the square root of the gas flow rate and dividing by the square root of the flow velocity. The flow correction coefficient is between 0.8 and 1.2. The gas flow rate is in cubic meters per second and the flow velocity is limited to no more than 25 meters per second.

[0053] The second pneumatic adjustment module 2 is used to add a second flexible pipeline between the secondary exhaust port and the aftercooler inlet. A second pneumatic adjustment valve is installed in the middle of the second flexible pipeline. The valve diameter is determined based on the diameter of the first pneumatic adjustment valve, specifically by multiplying the diameter of the first pneumatic adjustment valve by a coefficient ranging from 0.7 to 1.1.

[0054] Temperature and pressure monitoring module 3 is used to install an intake temperature sensor and an intake pressure sensor at the first-stage intake port, an exhaust temperature sensor and an exhaust pressure sensor at the first-stage exhaust port, and a total exhaust temperature sensor at the exhaust port of the whole machine.

[0055] PLC control module 4 is used to receive sensor data from the temperature and pressure monitoring module in real time, and to obtain the interstage compression ratio correction based on the difference between the actual exhaust temperature and the set temperature value and the deviation between the actual pressure and the designed 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 intake pressure and then by the designed compression ratio to obtain the deviation ratio), and finally multiplying the temperature difference by the temperature weighting factor and the pressure deviation value by the pressure weighting factor, and summing them up. The temperature weighting factor is between 0.35 and 0.55, and the pressure weighting factor is between 0.25 and 0.45.

[0056] The graded adjustment module 5 is used to dynamically adjust the opening of the first pneumatic control valve and the second pneumatic control valve respectively based on the interstage compression ratio correction under different load pressures.

[0057] The opening adjustment module 6 is used to generate valve opening commands based on the interstage compression ratio correction. The opening of the first pneumatic control valve is calculated based on the negative exponential function of the correction, and the opening of the second pneumatic control valve is calculated based on the positive exponential function of the correction. Specifically, the correction is multiplied by the valve characteristic constant and 100% after taking the negative exponent. The opening of the second pneumatic control valve is calculated based on the positive exponential function of the correction. Specifically, 1 is subtracted from the negative exponent of the correction and multiplied by the valve characteristic constant and 100%. The valve characteristic constant is between 0.8 and 1.5.

[0058] As described in modules 1-6 above, this invention constructs a collaborative control architecture that integrates flexible pipeline adjustment loops, multi-point monitoring, dynamic correction calculation, and graded opening control. Specifically, it sets up a first pneumatic adjustment module and a second pneumatic adjustment module to construct flexible pipelines between the intercooler outlet and the first-stage intake port, and the second-stage exhaust port and the aftercooler inlet, respectively. These are combined with a temperature and pressure monitoring module to collect temperature and pressure parameters of the first-stage intake, first-stage exhaust, and overall exhaust. The PLC control module calculates the interstage compression ratio correction based on the temperature difference and pressure deviation. The graded adjustment module adapts to different load pressure conditions. Finally, the opening control module regulates the opening of the two pneumatic adjustment valves according to an exponential function, achieving precise and flexible control of the pressure and temperature between the first and second stages. This solves the problem of excessively high first-stage exhaust temperature in oil-free screw air compressors under low-speed loads, which limits the minimum energy-saving speed. This achieves the goal of safe operation of the unit at lower speeds and further improves energy-saving performance.

[0059] The compression process of an oil-free screw air compressor relies on the meshing of the compressor rotors to achieve gas volume compression. Its compression efficiency is positively correlated with the rotational speed. When the unit is under low-speed load conditions, the reduced rotor speed leads to increased gas leakage in the rotor meshing gap, resulting in a decrease in compression efficiency. The gas cannot be fully compressed in the first stage of compression, and more of the input mechanical energy is converted into internal gas energy, directly manifesting as an increase in the first-stage exhaust temperature. Furthermore, traditional oil-free screw air compressors use fixed pipelines connecting the first and second stages of the compressor head, with the compression ratio determined during the design phase. This cannot adapt to the reduced gas flow velocity and changes in pipeline pressure loss at low speeds. Relying solely on the passive cooling function of the intercooler is insufficient to effectively control the first-stage exhaust temperature. If the rotational speed is further reduced to pursue energy savings, the first-stage exhaust temperature will exceed the limits of the compressor head seals and rotor clearance, potentially causing compressor head failure. This limits the minimum energy-saving speed of the unit, preventing it from fully realizing its energy-saving potential at low speeds.

[0060] Traditional interstage pressure and temperature control schemes for oil-free screw air compressors suffer from three core shortcomings: First, the pipeline design lacks flexibility, relying solely on an intercooler for passive cooling of the first-stage exhaust gas without an active adjustment loop. This makes it impossible to alter the intake and exhaust flow ratio between the first and second stage compressor heads, hindering the proactive adjustment of the interstage compression ratio. Second, the temperature and pressure monitoring dimensions are incomplete, focusing primarily on the overall exhaust temperature and pressure parameters while neglecting real-time monitoring of key measurement points such as the first-stage intake (affecting the initial conditions of first-stage compression) and the first-stage exhaust (directly reflecting the first-stage compression effect). This results in a one-sided control approach, failing to accurately assess dynamic changes in interstage operating conditions. Third, the control logic lacks precision, often employing simple on / off controls such as "starting the cooling fan when the temperature exceeds a threshold" and "shutting down the machine when the pressure exceeds a threshold." It fails to calculate the interstage compression ratio correction based on dynamic changes in temperature and pressure differences, and does not adjust valve openings according to different load pressure levels. This makes it difficult to adapt to the differences in operating conditions under low, medium, and high loads, resulting in poor control flexibility and adaptability. To address the aforementioned issues, this solution constructs active regulation capabilities by adding flexible pipelines and pneumatic regulating valves, supplements key measurement point monitoring to improve control basis, and combines PLC dynamic correction calculation and graded opening control to form a systematic solution covering the entire process, specifically solving the passive, one-sided, and low-precision problems of traditional solutions.

[0061] The core function of the first pneumatic regulation module is to construct a return flow regulation loop for the primary compressed air. This is achieved by adding a first flexible pipeline between the intercooler outlet (where the primary compressed air has been cooled to 40-60℃) and the primary intake port, and installing a first pneumatic regulating valve in the middle of this pipeline. This allows the cooled primary compressed air to be returned to the primary intake port. The returning cold air dilutes the intake temperature of the primary intake port and simultaneously regulates the intake flow rate of the primary compressor head, thereby actively changing the primary compression ratio. The valve diameter needs to be calculated based on the gas flow rate under the lowest load condition and the maximum allowable flow velocity of the pipeline. The specific formula is as follows:

[0062] ;

[0063] Where D represents the valve diameter and k represents the flow correction factor. This indicates the gas flow rate under the lowest load condition. Indicates the maximum permissible flow velocity in the pipeline;

[0064] The physical meaning of this calculation logic is as follows: Under the lowest load condition, the gas flow rate is minimal, and it must be ensured that the gas velocity in the pipeline does not exceed 25 meters per second (to avoid excessive pressure loss and turbulence noise due to excessive velocity, or slow adjustment response due to excessively low velocity). A flow correction coefficient of 0.8 to 1.2 is used to adapt to the different flow characteristics of different gas media (such as dry compressed air and compressed air containing trace impurities), ensuring that the diameter calculation results balance safety and adjustment flexibility. The required "gas flow rate under the lowest load condition" is obtained based on the linear correspondence between the air compressor's rated displacement and speed. For example, when the air compressor's rated speed is 1500 r / min, the rated displacement is 10 m³ / s. At the lowest load speed of 500 r / min (the lowest energy-saving speed in traditional solutions), the gas flow rate is converted proportionally to 3.33 m³ / s. The "maximum allowable flow velocity in the pipeline" is preset to 25 meters per second based on the pressure resistance of the pipeline material and the gas flow resistance characteristics. Taking specific parameters as an example, if the gas flow rate under the lowest load condition is 3.33 m³ / s, the flow correction factor is taken as 1.0, and the flow velocity is taken as 20 m / s (the limit below 25 m / s is reserved for safety margin), then the diameter of the first pneumatic regulating valve = 1.0 × ≈1.0×1.825÷4.472≈0.408 meters. This diameter ensures stable gas flow in the pipeline under minimum load, meeting the backflow regulation requirements while preventing the first-stage intake pressure from fluctuating beyond the safe range of ±0.02MPa due to excessive flow velocity.

[0065] The second pneumatic adjustment module forms a coordinated adjustment relationship with the first pneumatic adjustment module. By adding a second flexible pipeline between the secondary exhaust port (high-pressure gas after secondary compression, with a pressure of 0.8-1.2 MPa) and the aftercooler inlet (i.e., the pipeline after the primary exhaust and before entering the aftercooler, with a pressure of 0.3-0.5 MPa), and installing a second pneumatic adjustment valve in the middle of the pipeline, the high-pressure gas of the secondary exhaust can be replenished into the primary exhaust section. The replenished high-pressure gas can adjust the primary exhaust pressure, indirectly changing the intake pressure of the secondary compressor head, thereby coordinating and optimizing the secondary compression ratio. Its valve diameter is determined based on the diameter of the first pneumatic regulating valve and is calculated by multiplying it by a coefficient of 0.7 to 1.1. The value of this coefficient is set according to the design compression ratio ratio of the first and second stage compressor heads. Since the design compression ratio of the second stage compressor head is usually higher than that of the first stage (e.g., the design compression ratio of the first stage is 3.0 and the design compression ratio of the second stage is 4.0), the exhaust flow rate of the second stage and the return flow rate of the first stage need to be matched. The coefficient of 0.7 to 1.1 can ensure that the air supply volume of the second pipeline matches the return flow rate of the first pipeline, and avoid the pressure imbalance between stages caused by excessive flow in a single pipeline. Here, "first pneumatic regulating valve diameter" refers to the value calculated and determined in the aforementioned first pneumatic regulating module (e.g., 0.408 meters). If the coefficient is taken as 0.9 (to match the 3:4 ratio of the first and second stage design compression ratios), then the second pneumatic regulating valve diameter = 0.408 × 0.9 ≈ 0.367 meters. This diameter allows the supplementary air volume of the second stage exhaust to maintain a 1:1.2 matching ratio with the cold air flow of the first stage return flow, jointly regulating the first stage exhaust pressure and providing stable intake conditions for the second stage compressor head.

[0066] The temperature and pressure monitoring module is the sensing center of the entire control system. It achieves real-time parameter acquisition by installing sensors at key measuring points: An intake temperature sensor (monitoring the first-stage intake temperature, such as 37.5℃) and an intake pressure sensor (monitoring pressure typically 0.09-0.11MPa) are installed at the first-stage intake port 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 first-stage compression, directly reflecting the first-stage compression efficiency and heat accumulation) and an exhaust pressure sensor (monitoring the gas pressure after first-stage compression) are installed at the first-stage exhaust port to obtain the core parameters of the first-stage compression effect; and a total exhaust temperature sensor (monitoring the final output gas temperature, typically 40-50℃) is installed at the overall exhaust port to assist in judging the overall operating status of the unit. All analog signals (such as DC 4-20mA current signals) collected by the sensors are transmitted to the digital controller, converted into digital signals, and then transmitted to the PLC control module, providing real-time and comprehensive parameter data for subsequent correction calculations. For example, when the first-stage intake temperature sensor collects Ts1 as 37.5℃ and the first-stage exhaust temperature sensor collects a temperature of 135℃, the temperature rise of the first-stage compression can be directly obtained as 97.5℃, providing a core basis for determining whether to increase the cold air recirculation and reduce the first-stage exhaust temperature.

[0067] The PLC control module serves as the central calculation unit. Based on real-time data transmitted from the temperature and pressure monitoring module, it quantitatively calculates the interstage compression ratio correction. This correction is the core basis for subsequent valve opening control. The specific calculation process consists of three steps: First, the temperature difference is calculated as "Temperature Difference = Actual Exhaust Temperature - Set Temperature Value," where "Actual Exhaust Temperature" is the real-time value collected by the first-stage exhaust temperature sensor, and "Set Temperature Value" is the preset safe temperature limit for first-stage exhaust based on the compressor head's temperature resistance characteristics. This setting may lead to reduced rotor clearance, causing rotor contact and other malfunctions. Second, the pressure deviation is calculated as "Pressure Deviation = (First-Stage Exhaust Pressure ÷ First-Stage Intake Pressure) ÷ Design Compression Ratio," where "First-Stage Exhaust Pressure" and "First-Stage Intake Pressure" are the first-stage exhaust and first-stage intake pressures, respectively. The real-time values ​​collected by the air pressure sensor, the "design compression ratio" is the rated compression ratio preset based on structural parameters such as the number of teeth and meshing clearance of the first-stage compressor rotor (e.g., 3.0), and the pressure deviation value reflects the degree of deviation between the actual compression ratio and the design value; the third step is to calculate the interstage compression ratio correction, i.e., "correction amount = temperature difference × temperature weighting factor + pressure deviation value × pressure weighting factor", where the temperature weighting factor is 0.35 to 0.55 and the pressure weighting factor is 0.25 to 0.45. The temperature weighting factor is higher than the pressure weighting factor because excessively high first-stage exhaust temperature directly threatens the safe operation of the unit, and temperature needs to be controlled by adjusting the compression ratio first; the pressure weighting factor is used to take into account the interstage pressure stability and avoid excessive pressure deviation causing compressor load fluctuations to exceed the safe range of ±5%. For example, if the actual first-stage exhaust temperature is 130℃ (set temperature value 120℃, temperature difference 10℃), the temperature weighting factor is 0.4; the first-stage exhaust pressure is 0.6MPa, the first-stage intake pressure is 0.2MPa (actual compression ratio = 0.6 ÷ 0.2 = 3.0), the design compression ratio is 3.0 (pressure deviation = 3.0 ÷ 3.0 = 1.0, i.e. no deviation), and the pressure weighting factor is 0.3; then the interstage compression ratio correction is 10 × 0.4 + (1.0 - 1.0) × 0.3 = 4.0. This correction indicates that the first-stage exhaust temperature needs to be reduced by about 10℃ by increasing the first-stage cold air recirculation and adjusting the compression ratio to return it to the safe range.

[0068] The function of the staged adjustment module is to achieve operating condition adaptation. Based on the operating characteristics under different load pressures, it differentiates the opening of the two pneumatic regulating valves. "Different load pressures" are determined by the exhaust pressure collected by the exhaust pressure sensor. The specific classification criteria are: exhaust pressure ≤ 0.6MPa corresponds to low load (speed ≤ 800r / min), 0.6MPa < exhaust pressure ≤ 0.9MPa corresponds to medium load (speed 800-1200r / min), and exhaust pressure > 0.9MPa corresponds to high load (speed > 1200r / min). The "interstage compression ratio correction" is the calculation result of the PLC control module. The significance of this module is as follows: Under low load conditions, the first-stage compressor head has the lowest compression efficiency, requiring a focus on increasing the cold air return through the first pneumatic regulating valve to reduce the first-stage exhaust temperature; under medium load conditions, the compression efficiency and gas flow characteristics are relatively stable, requiring balanced adjustment of the two valves to maintain stable interstage parameters; under high load conditions, the compression load of the second-stage compressor head increases, requiring optimization of the supplementary air volume through the second pneumatic regulating valve to stabilize the second-stage compression ratio. For example, under low load pressure (total exhaust pressure 0.5MPa), the interstage compression ratio correction is 5.0. The stage adjustment module prioritizes increasing the opening of the first pneumatic regulating valve, while the second pneumatic regulating valve maintains a smaller opening. Under high load pressure (total exhaust pressure 1.0MPa), the correction is 2.0. The module prioritizes increasing the opening of the second pneumatic regulating valve, while the first pneumatic regulating valve maintains a basic opening, ensuring that the adjustment effect accurately matches the operating conditions.

[0069] The function of the valve opening adjustment module is to convert the interstage compression ratio correction into specific valve action commands. It achieves smooth valve opening control through an exponential function, avoiding drastic fluctuations in interstage pressure and temperature caused by sudden changes in opening. Specifically, the opening of the first pneumatic control valve is calculated based on a negative exponential function of the correction, with the formula: "First pneumatic control valve opening = Negative exponent of correction × Valve characteristic constant × 100%"; the opening of the second pneumatic control valve is calculated based on a positive exponential function of the correction, with the formula: "Second pneumatic control valve opening = (1 - Negative exponent of correction) × Valve characteristic constant × 100%". The "valve characteristic constant" ranges from 0.8 to 1.5, determined based on the flow characteristic curve of the pneumatic control valve (e.g., 1.0 for linear flow characteristic valves, 1.2 for equal percentage flow characteristic valves), to adapt to the flow opening response characteristics of different valves. The significance of this calculation logic is that the first pneumatic control valve controls the cold air return flow. The larger the interstage compression ratio correction (the larger the temperature and pressure deviation), the more cold air needs to return. The interstage compression ratio correction is represented by w, and the negative exponential function ( The first pneumatic regulating valve can achieve a gradual increase in the opening degree when the correction amount increases (e.g., when the correction amount increases from 4.0 to 8.0, e^-4.0≈0.018, e^-8.0≈0.0003, the opening degree gradually decreases from 1.8% to 0.03%. Here, it is necessary to combine the benchmark opening degree adjustment of the graded adjustment module to ensure that the actual opening degree is consistent with the correction requirement), avoiding a sudden increase in the opening degree that would cause a sudden drop in the first-stage intake pressure; the second pneumatic regulating valve controls the second-stage exhaust air supply. The larger the correction amount, the less the air supply needs to be. The positive exponential function can achieve a gradual decrease in the opening degree when the correction amount increases (e.g., when the correction amount increases from 4.0 to 8.0, the opening degree gradually increases from 98.2% to 99.97%), avoiding a sudden drop in the opening degree that would cause a sudden increase in the first-stage exhaust pressure. Taking a valve characteristic constant of 1.0 as an example, when the interstage compression ratio correction is 4.0, the opening of the first pneumatic regulating valve is approximately 1.8% (e^-4.0×1.0×100%), and the opening of the second pneumatic regulating valve is approximately 98.2% (1-e^-4.0)×1.0×100%). The two valves work together to adjust the first-stage compression ratio from 3.0 to 2.8, thus achieving a gradual reduction in the first-stage exhaust temperature.

[0070] Through the synergistic effect of the above six modules, the system can adapt to changes in operating conditions of the oil-free screw air compressor under different speeds and loads in real time, actively adjusting the pressure and temperature between the first and second stages. This effectively solves the problem of excessively high first-stage exhaust temperature at low speeds, allowing the minimum energy-saving speed of the unit to be lower than that of traditional solutions. Furthermore, the first-stage exhaust temperature can be controlled below the limit threshold during low-speed operation, thus improving the unit's energy-saving effect. This avoids the passivity and blindness of traditional solutions, significantly improving control precision and flexibility, and enabling stable adaptation to a full range of operating conditions with load rates from 20% to 100%.

[0071] In one embodiment of the present invention, the PLC control module includes:

[0072] The data acquisition unit is used to acquire the load rate of the air compressor in real time. The load rate of the air compressor is calculated by comparing the current speed with the rated speed.

[0073] The first calculation unit is used to calculate the load correction coefficient, which is determined based on the load rate segment; 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; when the load rate is greater than 70%, the coefficient is equal to 0.7 plus 0.2 times the load rate.

[0074] The first correction unit is used to correct the opening degree of the first pneumatic control valve by multiplying the initial opening degree by the load correction coefficient to obtain the corrected opening degree of the first pneumatic control valve.

[0075] The second correction unit is used to correct the opening degree of the second pneumatic control valve. The corrected opening degree of the second pneumatic control valve is obtained by multiplying the initial opening degree by 2 and subtracting the load correction coefficient.

[0076] The forced correction unit is used to forcibly increase the opening of the corrected first pneumatic regulating valve by 20% to 35% when the actual exhaust temperature exceeds the alarm threshold.

[0077] As described above, by constructing closed-loop control logic, the acquisition unit obtains the air compressor load rate based on the rotational speed, the first calculation unit determines the load correction coefficient segmented according to the load rate range, the first correction unit and the second correction unit respectively use different calculation methods to correct the initial opening of the first pneumatic regulating valve and the second pneumatic regulating valve, and the forced correction unit triggers a forced increase in opening when the actual exhaust temperature exceeds the alarm threshold, so as to achieve accurate adaptation of 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 changes dynamically, ensure that the unit can maintain stable interstage pressure and temperature in the low, medium and high full load rate range, further expand the load range for safe operation of the unit, and improve the system's adaptability to load fluctuations.

[0078] The load rate of an oil-free screw air compressor is directly related to the rotor speed of the compressor head, which determines the characteristics of the gas compression process within the compressor head. When the load rate decreases, the rotor speed decreases accordingly, increasing the proportion of gas leakage in the rotor meshing gap. During the first-stage compression process, the gas cannot be fully compressed, and more mechanical energy is converted into internal gas energy, leading to a more pronounced rise in the first-stage exhaust temperature. Simultaneously, at low load rates, the gas flow velocity in the pipeline decreases. If the initial opening of the first pneumatic regulating valve is calculated solely based on the interstage compression ratio correction, it may not provide sufficient cold air return flow to suppress the temperature rise. When the load rate increases, the rotor speed increases, the first-stage compression efficiency improves, and the gas compression volume increases. The intake volume and pressure of the second-stage compressor head rise synchronously. If the second pneumatic regulating valve still uses a fixed logic correction opening, it can easily cause the second-stage exhaust pressure fluctuation to exceed the design range (typically ±0.05MPa), increasing the operating load on the compressor head. Furthermore, when the actual exhaust temperature exceeds the alarm threshold of the conventional adjustment capability, the conventional valve opening correction speed cannot meet the rapid cooling requirements. If not intervened in time, it will lead to rotor expansion, narrowing of the rotor gap, and even rotor collision and jamming, causing shutdown failure. Therefore, it is necessary to dynamically adjust the valve opening based on the load rate and set up an over-temperature forced intervention mechanism to adapt to the compression characteristics under different loads and cope with extreme over-temperature conditions.

[0079] The core function of the acquisition unit is to obtain the load rate parameter, which reflects the actual load status of the air compressor. This parameter is calculated by "load rate = current speed ÷ rated speed × 100%". The "current speed" is acquired in real time by a speed sensor installed at the rotor shaft end of the air compressor head (acquisition frequency is 1 time / second, signal type is DC4-20mA, corresponding to a speed range of 0-30000r / min). The "rated speed" is the rated operating speed determined during the air compressor design stage based on the material strength and compression efficiency of the rotor head (e.g., 1500r / min, this value is a fixed design parameter and is stored in the parameter library of the PLC control module). The load rate directly reflects the actual operating load of the turbine rotor. For example, if the current speed is 450 r / min and the rated speed is 1500 r / min, then the load rate = 450 ÷ 1500 × 100% = 30%, indicating that the unit is operating under low load. The opening of the first pneumatic regulating valve needs to be adjusted to increase the return flow of cold air. If the current speed is 1200 r / min and the rated speed is 1500 r / min, then the load rate = 1200 ÷ 1500 × 100% = 80%, indicating that the unit is operating under high load. The opening of the second pneumatic regulating valve needs to be optimized to stabilize the secondary intake pressure. The load rate data acquired by the acquisition unit is transmitted to the first calculation unit in real time, providing a core basis for subsequent correction coefficient calculations. Its physical significance lies in converting the mechanical operating parameter of the turbine rotor speed into a quantifiable load index, providing a calculation benchmark that matches the actual operating conditions for the opening correction.

[0080] The function of the first calculation unit is to calculate the load correction coefficient based on the load rate range. The design logic of this coefficient is directly related to the unit's operating characteristics under different load rates. The specific segmentation rules and calculation methods are as follows: When the load rate is less than 40%, a linear calculation method of "load correction coefficient = 1.2 - 0.3 × load rate" is adopted. In this range, the unit is in a low-load state, the rotor speed is low, the gas leakage is large, and the first-stage exhaust temperature is easy to rise. It is necessary to increase the opening of the first pneumatic regulating valve through a larger correction coefficient to increase the cold air return flow. Therefore, the coefficient increases as the load rate decreases. For example, when the load rate is 30%, the coefficient = 1.2 - 0.3 × 0.3 = 1.11, and when the load rate is 20%, the coefficient = 1.2 - 0.3 × 0.2 = 1.14, ensuring that the lower the load rate, the larger the correction coefficient, and the more significant the increase in the opening of the first pneumatic regulating valve. When the load rate is between 40% and 70%, the load correction coefficient is fixed at 0.85. In this range, the unit is in a medium-load state. Under load conditions, with a moderate rotor speed (typically 600-1050 r / min), gas leakage and compression efficiency are in a stable balance. The first-stage exhaust temperature and second-stage intake pressure fluctuate little, eliminating the need for dynamic adjustment of the correction coefficient. The fixed coefficient ensures stable opening correction logic and avoids parameter fluctuations caused by frequent adjustments. When the load rate is greater than 70%, a linear calculation method of "load correction coefficient = 0.7 + 0.2 × load rate" is adopted. Within this range, the unit is under high load, with high rotor speed and large gas compression. The second-stage compressor head intake pressure increases synchronously with the load. It is necessary to optimize the opening of the second pneumatic regulating valve to stabilize the second-stage compression process. Therefore, the coefficient increases with the increase of the load rate. For example, when the load rate is 80%, the coefficient = 0.7 + 0.2 × 0.8 = 0.86, and when the load rate is 90%, the coefficient = 0.7 + 0.2 × 0.9 = 0.88. This ensures that the higher the load rate, the larger the correction coefficient, and the more suitable the second pneumatic regulating valve opening correction is for the high load requirements of the second stage. The load correction coefficients calculated by this 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 correction logic is accurately matched with the core requirements of the unit under low, medium and high loads, avoiding the adjustment inaccuracy caused by a single coefficient.

[0081] The function of the first correction unit is to correct the initial opening of the first pneumatic control valve. The correction method is "corrected opening of the first pneumatic control valve = initial opening of the first pneumatic control valve × load correction coefficient", where "initial opening of the first pneumatic control valve" is the opening value calculated by the opening adjustment module (calculated based on the negative exponential function of the interstage compression ratio correction). For example, if the initial opening of the first pneumatic control valve is 20%, the load rate obtained by the acquisition unit is 30%, and the load correction coefficient calculated by the first calculation unit is 1.11, then the corrected opening of the first pneumatic control valve = 20% × 1.11 = 22.2%. The significance of this correction logic lies in the following: The core function of the first pneumatic regulating valve is to control the cold air recirculation flow from the intercooler outlet to the first-stage intake port. Under low load, the load correction coefficient is greater than 1, which can increase the initial opening to increase the cold air recirculation, directly reducing the intake temperature of the first-stage intake port, reducing the internal energy conversion during the first-stage compression process, and thus suppressing the rise in the first-stage exhaust temperature. Under medium load, the coefficient is 0.85 (less than 1), which moderately reduces the initial opening to avoid excessive recirculation leading to excessively low first-stage intake pressure (below 0.08 MPa). Under high load, the coefficient increases with the load rate, which can maintain or slightly increase the opening to balance the first-stage compression efficiency and temperature control requirements. For example, at a load rate of 80%, the initial opening of the first pneumatic regulating valve is 15%, and the load correction coefficient is 0.86. The corrected opening = 15% × 0.86 = 12.9%. This opening can ensure the stability of the first-stage exhaust temperature while avoiding excessive recirculation affecting the first-stage compression, thus ensuring the intake requirements of the second-stage compressor head.

[0082] 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)". The "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 interstage compression ratio correction). 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. When the load is low, the load correction coefficient is larger and the "2 - load correction coefficient" is smaller, which can reduce the opening of the second pneumatic control valve, reduce the secondary exhaust air supply, and avoid excessively high primary exhaust pressure. When the load is high, the load correction coefficient is larger and the "2 - load correction coefficient" is also larger, which can increase the opening of the second pneumatic control valve and increase the air supply to stabilize the secondary intake pressure. For example, if the initial opening of the second pneumatic control valve is calculated to be 30%, and the load rate is 30%, the load correction factor is 1.11. "2 - load correction factor" = 2 - 1.11 = 0.89, and the corrected opening is 30% × 0.89 = 26.7%. This opening reduces the amount of supplementary air at low loads, preventing the first-stage exhaust pressure from exceeding the design range of 0.3-0.5 MPa. If the load rate is 80%, the load correction factor is 0.86. "2 - load correction factor" = 2 - 0.86 = 1.14, and the corrected opening is 30% × 1.14 = 34.2%. This opening increases the amount of supplementary air at high loads, ensuring the second-stage intake pressure remains stable at 0.6-0.8 MPa, matching the high compression requirements of the second stage. The physical significance of this unit lies in using a differentiated correction formula to make the opening of the second pneumatic control valve and the first pneumatic control valve synergistic and complementary, jointly maintaining interstage pressure balance and avoiding parameter imbalances caused by single valve adjustment.

[0083] The forced correction unit is designed to handle extreme over-temperature conditions. When the actual exhaust temperature collected by the primary exhaust temperature sensor in the temperature and pressure monitoring module exceeds the alarm threshold (this threshold is determined based on the highest temperature resistance of the compressor head seal; depending on the compressor head design, the exhaust temperature may reach 250-300℃ or even higher), the forced correction logic is triggered, increasing the opening of the first pneumatic regulating valve by 20% to 35%. For example, if the corrected opening of the first pneumatic regulating valve is 22.2%, and the actual exhaust temperature reaches 165℃ (exceeding the 160℃ alarm threshold), a forced increase of 30% will result in a final opening of 22.2% × (1 + 30%) = 28.86%. The physical significance of this unit lies in the following: when the actual exhaust temperature exceeds the adjustment capability of the conventional correction logic, the opening of the first pneumatic regulating valve is forcibly increased, significantly increasing the cold air return flow rate and achieving rapid cooling. The opening adjustment rate of the conventional correction logic is 1%-2% per second, while the forced correction can complete a 20%-35% opening increase within 1 second, shortening the cooling response time to 5-10 seconds, which is much faster than conventional regulation. This effectively avoids the shutdown protection threshold caused by further increases in the actual exhaust temperature. For example, if the actual exhaust temperature starts at 165 degrees Celsius, after forcibly increasing the opening, the cold air return flow rate increases by 30%, the intake temperature at the first-stage intake port drops from 35 degrees Celsius to 30 degrees Celsius, and the first-stage exhaust temperature drops to 145 degrees Celsius within 8 seconds, returning to the controllable range of conventional regulation, thus preventing production interruptions caused by over-temperature shutdown.

[0084] Through the synergistic effect of the aforementioned units, load adaptation optimization of the basic control logic is achieved: In the low load range, the opening of the first pneumatic regulating valve is increased by a larger load correction coefficient, effectively suppressing the rise in primary exhaust temperature, enabling 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, a fixed correction coefficient ensures parameter stability, with primary exhaust temperature fluctuations controlled within ±5 degrees Celsius and secondary intake pressure fluctuations controlled within ±0.02 MPa. In the high load range, the overshoot of the secondary exhaust pressure is reduced by optimizing the opening of the second pneumatic regulating valve. Under over-temperature conditions, the forced correction mechanism shortens the cooling response time, effectively avoiding over-temperature shutdown. Overall, through dynamic load rate adaptation and emergency intervention, the control accuracy and stability of the system across the entire load range are further improved, providing more comprehensive protection for the safe and energy-saving operation of the unit.

[0085] In one embodiment of the present invention, the temperature and pressure monitoring module includes:

[0086] The detection unit is used to monitor the rate of change of the primary exhaust temperature, which is calculated based on the temperature change value per unit time, in degrees Celsius per second.

[0087] The second calculation unit is used to calculate the thermal inertia compensation coefficient. The thermal inertia compensation coefficient is determined based on the temperature change rate. First, the temperature change rate is divided by the reference change rate (1.5 to 3.0 degrees Celsius per second). Then, the hyperbolic tangent function value is taken, multiplied by the compensation intensity factor (0.1 to 0.3), and finally 1 is added to obtain the coefficient value.

[0088] The valve opening adjustment unit is used to apply the thermal inertia compensation coefficient to the valve opening adjustment. The opening of the first pneumatic regulating valve is calculated by multiplying the corrected opening of the first pneumatic regulating valve by the compensation coefficient. The opening of the second pneumatic regulating valve is calculated by multiplying the corrected opening of the second pneumatic regulating valve by 2 and subtracting the difference from the thermal inertia compensation coefficient.

[0089] The emergency cooling unit is used to trigger the emergency cooling mode when the temperature change rate is greater than the preset value (the preset value is 5 degrees Celsius per second). Ultimately, the opening of the first pneumatic regulating valve increases by 40%, and the opening of the second pneumatic regulating valve decreases by 50%.

[0090] As described above, the unit-time change rate of the first-stage exhaust temperature is obtained by the detection unit. The second calculation unit calculates the thermal inertia compensation coefficient based on this change rate, combined with the hyperbolic tangent function and the compensation intensity factor. The opening adjustment unit applies the compensation coefficient to the corrected opening of the first and second pneumatic regulating valves to offset the thermal inertia effect. The emergency cooling unit triggers extreme condition intervention when the temperature change rate exceeds the preset value. This achieves targeted solutions to the two core problems of thermal inertia lag and rapid temperature change runaway in the interstage temperature control of the oilless screw air compressor, ensuring the timeliness and stability of the system's adjustment response during dynamic temperature changes, and avoiding overshoot caused by thermal inertia or temperature runaway caused by rapid temperature changes.

[0091] The interstage temperature regulation of oil-free screw air compressors exhibits significant thermal inertia. The change in the first-stage exhaust temperature depends not only on the cold air return flow of the first pneumatic regulating valve but also on the heat capacity of components such as the intercooler, connecting pipelines, and compressor housing. When the opening of the first pneumatic regulating valve increases the cold air return flow, the cold air must first exchange heat with the inner wall of the pipeline and the heat exchange elements of the intercooler before gradually reducing the overall gas temperature in the pipeline. This is reflected in the reading change of the first-stage exhaust temperature sensor with a time lag (usually 3-8 seconds). This lag is the regulation delay caused by thermal inertia. If the temperature regulation is based only on the instantaneous temperature value without considering the rate of temperature change, over-regulation is likely to occur. For example, when the first-stage exhaust temperature rises from 120℃ at a rate of 2℃ / s, conventional regulation requires waiting for the temperature sensor reading to exceed the set value (e.g., 125℃) before increasing the valve opening. However, by this time, thermal inertia has already caused the actual gas temperature in the pipeline to continue to rise, and the temperature may have exceeded 130℃ by the time the regulation takes effect, resulting in overshoot. Furthermore, when the unit encounters sudden operating conditions (such as a sudden increase in load from 50% to 80%), the rate of change of the primary exhaust temperature will increase dramatically (potentially exceeding 5°C / s). The response speed of conventional thermal inertia compensation regulation cannot meet the rapid cooling requirements. If strong intervention measures are not taken in time, the temperature will exceed the alarm threshold within seconds, threatening equipment safety. Therefore, it is necessary to calculate the thermal inertia compensation coefficient by monitoring the rate of temperature change, correct the valve opening in advance to offset the lag, and set an emergency cooling mode to deal with ultra-rapid temperature changes.

[0092] The core function of the detection unit is to acquire the rate of change 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 / second to ensure timely capture of dynamic temperature changes). The "time interval" is the time difference between two data acquisitions by the sensor (fixed to 1 second, this interval ensures data real-time performance while avoiding increased data noise due to excessively short intervals). The unit of the calculation result is degrees Celsius per second. The rate of temperature change directly reflects the dynamic trend of the first-stage exhaust temperature. For example, if the current first-stage exhaust temperature is 122 degrees Celsius, and the temperature was 120 degrees Celsius one second ago, then the rate of temperature change = (122-120) ÷ 1 = 2.0 degrees Celsius per second, indicating that the temperature is slowly rising. If the current temperature is 125 degrees Celsius, and the temperature was 120 degrees Celsius one second ago, then the rate of temperature change = 5.0 degrees Celsius per second, indicating that the temperature is rapidly rising. The temperature change rate data collected by this unit is transmitted to the second calculation unit in real time. Its significance lies in transforming the static instantaneous value of temperature into a dynamic trend value, providing a core basis for the subsequent calculation of the thermal inertia compensation coefficient that reflects the rate of temperature change, and avoiding the lag problem of traditional adjustment that only relies on instantaneous values.

[0093] 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 intensity by quantifying the temperature change rate. The larger the temperature change rate, the more significant the adjustment lag caused by thermal inertia, requiring a larger compensation coefficient to enhance the opening adjustment force. The smaller the temperature change rate, the weaker the thermal inertia effect, requiring a smaller compensation coefficient to avoid over-adjustment. The specific calculation process consists of four steps: First, 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 (taken as 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 value differences under different operating conditions. Second, 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 when the input value exceeds 1, the output value changes linearly. When the output value is outside this range, it tends to be ±1. This characteristic 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 change rate. When the change rate is far beyond the reference value, the compensation coefficient tends to stabilize at its maximum value, preventing over-compensation. The third step is to multiply the hyperbolic tangent function value by the compensation intensity factor (with a value of 0.1 to 0.3, for example, set to 0.2. The compensation intensity factor determines the maximum adjustment range of the compensation coefficient, avoiding excessive or insufficient compensation). The fourth step is to add 1 to the above product result to obtain the final thermal inertia compensation coefficient (the coefficient value is always greater than 1 to ensure that the compensation direction is to enhance the opening adjustment force). 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 strength factor is 0.2, the calculation process is as follows: 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 the compensation coefficient decreases synchronously when the temperature change rate decreases, achieving a precise match between the compensation strength and the temperature change rate. The thermal inertia compensation coefficient calculated by this unit is transmitted to the opening adjustment unit. Its significance lies in transforming the dynamic temperature change trend into a quantifiable basis for opening compensation, and achieving smooth adjustment of the compensation intensity through mathematical functions, thus avoiding the rigidity problem of traditional fixed compensation.

[0094] The function of the opening adjustment unit is to apply the thermal inertia compensation coefficient to correct the opening of the first and second pneumatic control valves. It employs differentiated correction logic based on the functional differences between the two valves. The first pneumatic control valve controls the cold air return flow, and its core function is to suppress the rise in first-stage exhaust temperature. Therefore, its opening needs to increase with the increase of the compensation coefficient. Thus, the calculation method is "first pneumatic control valve opening = first pneumatic control valve corrected opening × thermal inertia compensation coefficient," where "first pneumatic control valve corrected opening" is the opening value calculated by the first correction unit (already adjusted for load rate). The second pneumatic control valve controls the second-stage exhaust replenishment volume; increasing its opening will indirectly increase the first-stage exhaust volume. When the temperature changes rapidly, the opening of the pressure valve needs to be appropriately reduced to avoid pressure-assisted heating. Therefore, the calculation method of "second pneumatic control valve opening = second pneumatic control valve corrected opening × (2 - thermal inertia compensation coefficient)" is adopted. Here, "second pneumatic control valve corrected opening" is the opening value calculated by the second correction unit, and "2 - thermal inertia compensation coefficient" ensures that when the compensation coefficient increases, the second pneumatic control valve opening correction coefficient decreases synchronously (for example, when the compensation coefficient = 1.1523, 2 - 1.1523 = 0.8477; when the compensation coefficient = 1.0924, 2 - 1.0924 = 0.9076), so as to achieve reverse coordinated regulation of the two valves. For example, if the calculated corrected opening degree of the first pneumatic control valve is 22.2%, the corrected opening degree of the second pneumatic control valve is 26.7%, and the thermal inertia compensation coefficient is 1.1523, then the opening degree of the first pneumatic control valve = 22.2% × 1.1523 ≈ 25.6%, which is 3.4 percentage points higher than the corrected opening degree, increasing the cold air return flow to offset the thermal inertia lag; the opening degree of the second pneumatic control valve = 26.7% × 0.8477 ≈ 22.6%, which is 4.1 percentage points lower than the corrected opening degree, reducing the amount of supplementary air to avoid pressure-assisted heating. The significance of this correction logic lies in the fact that by using dual valves to adjust in reverse and in a coordinated manner, it not only enhances the active intervention of temperature control but also maintains the interstage pressure basically stable (pressure fluctuations can be controlled within ±0.02MPa), avoiding parameter imbalances caused by adjusting a single valve. For example, when adjusting only the first valve in the traditional way, the pressure fluctuation can reach ±0.05MPa, while this solution reduces the pressure fluctuation amplitude by 56% through synchronous correction using the second valve, achieving coordinated stability of temperature and pressure.

[0095] The emergency cooling unit is designed to handle extreme and rapid temperature changes. When the rate of change of the primary exhaust temperature monitored by the detection unit exceeds 5 degrees Celsius per second (this preset value is determined based on the temperature resistance limit of the head material and the sensor response speed; a rate of change of 5 degrees Celsius per second means that without intervention, the temperature will rise from 120 degrees Celsius to 170 degrees Celsius within 10 seconds, exceeding the alarm threshold), the emergency cooling mode is triggered, forcibly adjusting the opening of two valves: the opening of the first pneumatic regulating valve is increased by 40% based on the value calculated by the opening adjustment unit to maximize the cold air return flow and achieve rapid cooling; the opening of the second pneumatic regulating valve is decreased by 50% based on the value calculated by the opening adjustment unit to minimize the amount of supplementary air and avoid pressure increase exacerbating the temperature rise. For example, if the opening of the first pneumatic regulating valve calculated by the opening adjustment unit is 25.6% and the opening of the second pneumatic regulating valve is 22.6%, after triggering the emergency cooling mode, the opening of the first pneumatic regulating valve = 25.6% × (1 + 40%) = 35.84%, the cold air return flow rate increases by 40% compared with the conventional compensation adjustment, and the inlet air temperature of the first-stage intake port can drop from 35 degrees Celsius to 28 degrees Celsius within 3 seconds; the opening of the second pneumatic regulating valve = 22.6% × (1 - 50%) = 11.3%, the air supply is reduced by 50%, and the first-stage exhaust pressure drops from 0.4 MPa to 0.35 MPa, avoiding pressure-assisted heating. When the rate of temperature change exceeds the adjustment capability of thermal inertia compensation, the vicious cycle of temperature rise-thermal inertia lag-adjustment failure is broken by extreme opening adjustment, shortening the cooling response time from 8-12 seconds of conventional compensation to 3-5 seconds, ensuring effective control before the temperature exceeds the safety threshold. After triggering the emergency cooling mode, the primary exhaust temperature can be reduced from 130 degrees Celsius (change rate of 5.5 degrees Celsius per second) to 120 degrees Celsius in just 4 seconds, effectively preventing the equipment from shutting down due to overheating.

[0096] Through the synergistic effect of the aforementioned units, refined control of dynamic temperature changes between stages is achieved: In scenarios with normal temperature changes (rate of change ≤ 3 degrees Celsius per second), the thermal inertia compensation coefficient reduces the overshoot of the first-stage exhaust temperature regulation and improves regulation stability; in scenarios with moderately rapid changes (3 < rate of change ≤ 5 degrees Celsius per second), the compensation coefficient and valve coordinated correction shorten the temperature response time, preventing the temperature from approaching the alarm threshold; in scenarios with extremely rapid changes (rate of change > 5 degrees Celsius per second), the emergency cooling mode can curb the temperature rise within 5 seconds, reducing the number of equipment shutdowns due to over-temperature compared to traditional solutions. Overall, by monitoring the rate of temperature change and dynamic compensation, the regulation lag and overshoot problems caused by thermal inertia are solved. Emergency intervention to deal with extreme temperature changes further improves the system's temperature control capabilities under dynamic operating conditions, providing more comprehensive temperature protection for the safe and stable operation of the unit.

[0097] In one embodiment of the present invention, the PLC control module further includes:

[0098] The preset unit is used to preset the first-level 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;

[0099] The first opening adjustment unit is used to adjust the opening of the first pneumatic regulating valve when the actual first-stage exhaust pressure is less than the minimum safe pressure; the opening is calculated by multiplying the final opening by the actual pressure and dividing by the minimum safe pressure.

[0100] The second opening adjustment unit is used to adjust the opening of the second pneumatic regulating valve when the actual first-stage exhaust pressure is greater than the maximum safe pressure: the opening is calculated by adding (final opening minus 80%) multiplied by the difference between the actual pressure and the maximum safe pressure and dividing by the maximum safe pressure.

[0101] The third opening adjustment unit is used to close the second pneumatic regulating valve and fully open the first pneumatic regulating valve when the actual exhaust temperature exceeds the set threshold.

[0102] As described above, by constructing a pressure closed-loop management system, the minimum and maximum safe boundaries of the first-stage exhaust pressure are defined by a preset unit. When the pressure is lower than the minimum safe value, the first opening adjustment unit adjusts the opening of the first pneumatic regulating valve accordingly. When the pressure is higher than the maximum safe value, the second opening adjustment unit adjusts the opening of the second pneumatic regulating valve differently. When the actual exhaust temperature exceeds the limit threshold, the third opening adjustment unit triggers extreme valve action, thereby achieving full-condition safety control of the first-stage exhaust pressure. This solves the problems of insufficient response accuracy and insufficient intervention in extreme over-temperature conditions in the basic adjustment logic mentioned above, ensuring that the first-stage exhaust pressure is always within the safe operating range. At the same time, it provides a final safety guarantee for extreme over-temperature conditions, further improving the safety and stability of the system operation.

[0103] The primary discharge pressure is a core parameter connecting the primary and secondary compression processes of an oil-free screw air compressor. Its value directly determines the synergistic efficiency of the two stages and equipment safety. When the primary discharge pressure is lower than the minimum safe pressure, the intake pressure of the secondary compressor head is insufficient, leading to a reduction in gas compression during the secondary compression process. The overall discharge volume is lower than the design value (usually reduced by 15%-25%), failing to meet downstream gas demand. When the primary discharge pressure is higher than the maximum safe pressure, the compression load of the primary compressor head exceeds its design capacity, increasing the force at the rotor meshing point. This can easily lead to accelerated bearing wear and excessive compressor head vibration (exceeding 0.15 mm / s), shortening the compressor head's service life over long-term operation. Conventional valve opening adjustments cannot quickly cool the compressor; extreme valve actions are needed to cut off the heat source and maximize cooling to avoid irreversible failures such as seal melting and lubrication failure. Therefore, it is necessary to define the control boundary by pre-setting a safe range for the primary discharge pressure, design differentiated adjustment logic for different pressure anomaly scenarios, and set up an over-temperature extreme intervention mechanism to ensure the coordinated operation of the two stages and equipment safety.

[0104] The core function of the preset unit is to define the safe operating boundary of the primary exhaust pressure, that is, to preset the safe range of the primary exhaust pressure. The "design pressure" is the rated primary exhaust pressure determined during the design phase of the oil-free screw air compressor's primary head based on structural parameters such as the number of rotor teeth, meshing clearance, and rated speed (e.g., 0.4 MPa; the rated primary exhaust pressure is a fixed design parameter stored in the PLC control module's parameter library, directly determining the rated output of the primary compressor and the design intake conditions of the secondary head). The "minimum safe pressure = design pressure × 0.7" value, which is 0.7 times the design pressure, physically ensures that the secondary head receives the minimum intake pressure. If the primary exhaust pressure is lower than this value, the secondary compressor head's intake pressure will be insufficient, and the gas filling rate in the rotor meshing gap will be less than 80%, resulting in the secondary compression volume failing to meet design requirements and a significant decrease in the overall machine's exhaust volume. "Maximum safe pressure = design pressure × 1.3" means that the physical meaning of 1.3 times the design pressure is to protect the primary compressor head from overload damage. If the primary exhaust pressure is higher than this value, the compression reaction force borne by the primary compressor head rotor will exceed the design load limit (usually the design limit is 1.3 times the rated load), the bearing wear rate will accelerate by 3-5 times, and the compressor head vibration amplitude will exceed the safety limit of 0.15 mm / s. Taking a design pressure of 0.4 MPa as an example, the preset safe range for primary exhaust pressure is 0.28 MPa (0.4 × 0.7) to 0.52 MPa (0.4 × 1.3). This range ensures both the secondary compressor head's intake volume and the primary compressor head's safety, providing a clear control benchmark for subsequent pressure regulation.

[0105] The function of the first opening adjustment unit is to deal with scenarios where the first-stage exhaust pressure is lower than the minimum safe pressure. It increases the pressure by adjusting the opening of the first pneumatic regulating valve. The adjustment logic is "adjusted opening of the first pneumatic regulating valve = final opening of the first pneumatic regulating valve × (actual first-stage exhaust pressure ÷ minimum safe pressure)". The "actual first-stage exhaust pressure" comes from the exhaust pressure sensor installed at the first-stage exhaust port in the temperature and pressure monitoring module (collecting data once per second to ensure real-time capture of pressure dynamics). The "minimum safe pressure" is a value determined by the preset unit (e.g., 0.28 MPa). The "final opening of the first pneumatic regulating valve" is the corrected opening value (which incorporates load rate and thermal inertia compensation, e.g., 25.6%). The first pneumatic regulating valve controls the cold air return flow from the intercooler outlet to the first-stage intake port. When the first-stage exhaust pressure is lower than the minimum safe pressure, the cold air return flow needs to be reduced to increase the intake pressure of the first-stage intake port, thereby increasing the exhaust pressure of the first-stage compressor. Through the proportional coefficient of "actual pressure ÷ minimum safe pressure", the opening adjustment range and the pressure loss degree are precisely matched. The greater the pressure loss, the smaller the proportional coefficient, the greater the reduction in opening, the more significant the reduction in return flow, and the faster the pressure rise. For example, if the final opening of the first pneumatic regulating valve is 25.6%, and the actual first-stage exhaust pressure is 0.25 MPa (lower than the minimum safe pressure of 0.28 MPa), then the adjusted opening = 25.6% × (0.25 ÷ 0.28) ≈ 25.6% × 0.89 ≈ 22.8%, which is 2.8 percentage points lower than the final opening. The cold air return flow is reduced by 2.8%, the first-stage intake pressure increases from 0.09 MPa to 0.11 MPa, and the first-stage compression exhaust pressure rises from 0.25 MPa to 0.29 MPa within 5 seconds, returning to the safe range. This effectively solves the problem of insufficient intake of the second-stage compressor head, and the overall exhaust volume recovers from 82% of the design value to 98%.

[0106] The function of the second opening adjustment unit is to deal with scenarios where the primary exhaust pressure is higher than the maximum safe pressure. It reduces the pressure by adjusting the opening of the second pneumatic regulating valve. The adjustment logic is: "Adjusted opening of the second pneumatic regulating valve = final opening of the second pneumatic regulating valve + (final opening of the second pneumatic regulating valve - 80%) × (actual primary exhaust pressure - maximum safe pressure) ÷ maximum safe pressure". Here, "final opening of the second pneumatic regulating valve" is also the corrected opening value (e.g., 22.6%), "actual primary exhaust pressure" is the value collected by the exhaust pressure sensor (e.g., 0.55MPa), and "maximum safe pressure" is the value determined by the preset unit (e.g., 0.52MPa).

[0107] The design logic of this formula needs to be broken down and explained: "(Actual first-stage exhaust pressure - maximum safe pressure) ÷ maximum safe pressure" is the pressure over-limit ratio, quantifying the degree to which the pressure exceeds the safe boundary; "(Final opening of the second pneumatic regulating valve - 80%)" is the opening benchmark correction term. When the final opening is higher than 80%, this value is positive, which can increase the adjustment range and avoid the pressure from continuously rising under high opening; when the final opening is lower than 80%, this value is negative, which can reduce the adjustment range and prevent the pressure from dropping sharply due to excessive reduction of the opening; the product of the two is added to the final opening to achieve a two-factor dynamic adjustment of "pressure over-limit degree + current opening status", ensuring that the opening adjustment effectively reduces the pressure while avoiding excessive pressure fluctuations. For example, if the final opening of the second pneumatic regulating valve is 22.6%, and the actual first-stage exhaust pressure is 0.55 MPa (exceeding the maximum safe pressure of 0.52 MPa), then the pressure over-limit ratio = (0.55 - 0.52) ÷ 0.52 ≈ 0.058, the opening reference correction item = 22.6% - 80% = -57.4%, and the adjusted opening = 22.6% + (-57.4%) × 0.058 ≈ 22.6% - 3.33% ≈ 19.27%, which is 3.33 percentage points lower than the final opening. The second-stage exhaust air supply is reduced by 3.33%, and the first-stage exhaust pressure drops from 0.55 MPa to 0.51 MPa within 6 seconds, returning to the safe range. At the same time, the first-stage head load drops from 135% of the rated load to 128%, and the vibration amplitude drops from 0.18 mm / s to 0.14 mm / s, meeting the requirements for safe operation.

[0108] The function of the third opening adjustment unit is to cope with extreme overheating conditions where the actual exhaust temperature exceeds 180 degrees Celsius. At this time, conventional opening adjustment can no longer meet the rapid cooling requirements, and extreme valve actions need to be triggered: "closing the second pneumatic regulating valve" and "fully opening the first pneumatic regulating valve". The "actual exhaust temperature" comes from the exhaust temperature sensor at the first-stage exhaust port. The physical meaning of "closing the second pneumatic regulating valve" is to cut off the air supply channel from the second-stage exhaust to the first-stage exhaust port, avoiding the continuous increase of air supply leading to an increase in the first-stage exhaust pressure, which would further aggravate the temperature rise. The physical meaning of "fully opening the first pneumatic regulating valve" is to maximize the cold air return flow from the intercooler outlet to the first-stage intake port (the return flow is the design maximum when the opening is 100%), diluting the intake temperature of the first-stage intake port with a large amount of cold air, directly reducing the internal energy conversion in the first-stage compression process, and achieving rapid cooling. For example, when the actual exhaust temperature reaches 185 degrees Celsius, the third opening adjustment unit immediately triggers an action: the second pneumatic regulating valve quickly closes from 19.27% ​​to 0% (closing time ≤ 1 second), completely cutting off the replenishment air volume; the first pneumatic regulating valve quickly opens from 22.8% to 100% (opening time ≤ 1 second), increasing the cold air return flow from the design value of 22.8% to 100%, reducing the intake air temperature of the first-stage intake port from 35 degrees Celsius to 25 degrees Celsius, and reducing the first-stage exhaust temperature from 185 degrees Celsius to 160 degrees Celsius within 4 seconds and to 140 degrees Celsius within 8 seconds. This effectively avoids the melting of seals and lubrication failure, reducing production interruption time compared to the traditional "over-temperature shutdown" solution (traditional shutdown requires more than 30 minutes to restart, while this solution can be resumed without shutdown), and significantly improving production continuity.

[0109] Through the synergistic effect of the aforementioned units, comprehensive safety control of the primary exhaust pressure is achieved across all scenarios: In scenarios with normal pressure fluctuations (0.28-0.52 MPa), the first and second opening adjustment units can control the primary exhaust pressure fluctuation within ±0.02 MPa, a 60% reduction compared to traditional solutions (±0.05 MPa); in scenarios with pressure exceeding limits (below 0.28 MPa or above 0.52 MPa), the adjustment response time is shortened to 5-6 seconds, an improvement compared to traditional solutions (15-20 seconds), effectively preventing equipment damage caused by continuous pressure anomalies; in extreme over-temperature scenarios (above 180 degrees Celsius), the third opening adjustment unit can reduce the temperature to a safe range within 8 seconds, reducing the number of equipment shutdowns due to over-temperature conditions compared to traditional solutions. Overall, by clearly defining pressure safety boundaries, providing precise adjustment for different scenarios, and intervening in extreme operating conditions, the system's pressure safety control capabilities are improved, providing crucial assurance for the safe and stable operation of oil-free screw air compressors under complex pressure conditions.

[0110] In one embodiment of the present invention, the graded adjustment module includes:

[0111] The low-load regulating unit is used to set the target compression ratio to 0.8 times the design compression ratio in low-load mode (load rate less than 40%), and increase the reference opening of the first pneumatic regulating valve to the corresponding range (the corresponding range is 60% to 80%).

[0112] The medium load adjustment unit is used to set the target compression ratio to the design compression ratio in medium load mode (load rate between 40% and 70%), and to enable proportional-integral-derivative closed-loop control. The opening adjustment amount is calculated through the closed-loop control.

[0113] It is calculated by multiplying the proportional coefficient by the interstage compression ratio correction and then integrating the integral coefficient by the interstage compression ratio correction.

[0114] The high-load regulating unit is used to set the target compression ratio to 1.2 times the design compression ratio in high-load mode (load rate greater than 70%), and to increase the reference opening of the second pneumatic regulating valve to the corresponding range (corresponding range is 70% to 90%).

[0115] The mode switching unit is used to employ a gradual transition method during mode switching, with the opening change rate limited to an increase or decrease of no more than 5% per second.

[0116] As described above, by constructing a full-load regulation architecture, the low-load regulation unit, medium-load regulation unit, and high-load regulation unit respectively set the corresponding target compression ratio and valve reference opening range for operating conditions with load rates less than 40%, 40% to 70%, and greater than 70%. In the medium-load operating condition, proportional-integral-derivative (PID) closed-loop control is used to optimize the opening adjustment amount. The mode switching unit adopts a gradual opening transition method to avoid sudden parameter changes. This achieves precise adaptation and regulation of interstage pressure and temperature of the oil-free screw air compressor in the low, medium, and high full load range. It solves the problems of insufficient regulation accuracy and mode switching fluctuation caused by the fixed target compression ratio and single valve reference opening of the basic stage regulation logic under different load conditions, ensuring that the unit can maintain high compression efficiency and stable interstage parameters in the full load operating range.

[0117] The load rate of an oil-free screw air compressor directly determines the rotor speed of the compressor head, and the difference in speed 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 below 800 r / min (taking the rated speed of 1500 r / min as an example), the proportion of gas leakage in the rotor meshing gap increases to 15%-20%, and the gas cannot be fully compressed in the first stage of compression. If the design compression ratio (such as 3.0) is maintained, more of the mechanical energy input to the first stage compressor head will be converted into the internal energy of the gas, and the first stage exhaust temperature is likely to exceed the safety threshold of 120℃. At the same time, the gas flow velocity in the pipeline decreases under low load, and the reference opening of the first pneumatic regulating valve needs to be increased to increase the cold air return flow in order to effectively suppress the temperature rise. When the load rate is between 40% and 70%, the rotor speed stabilizes at 800-1050 r / 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 fluctuation of interstage parameters mainly comes from small changes in downstream gas load. It is necessary to use PID closed-loop control to correct the opening adjustment in real time to maintain the stability of interstage pressure and temperature. When the load rate is greater than 70%, the rotor speed is higher than 1050 r / min, the first-stage compression increases significantly, and the intake volume and intake pressure demand of the second-stage compressor head increase simultaneously. If the design compression ratio is still used, it will lead to insufficient intake pressure of the second-stage compressor head, and the overall exhaust pressure will be lower than the design value (e.g., the design value of 0.8 MPa may drop to 0.7 MPa). 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 regulating valve to optimize the second-stage supplementary air volume to meet the high compression demand of the second stage. In addition, when switching between different load modes, if the valve opening changes abruptly, it will cause the gas flow and pressure in the pipeline to fluctuate instantaneously, with an amplitude of ±0.08MPa or ±10℃, which exceeds the safe tolerance range of the equipment. Therefore, it is necessary to control the opening change rate through a gradual transition method.

[0118] The function of the low-load regulating unit is to adapt to working conditions with a load rate of less than 40%. The core is to reduce the burden on the first stage compressor head and suppress temperature rise by reducing the target compression ratio and increasing the reference opening of the first pneumatic regulating valve. The "load rate" is calculated by the acquisition unit (load rate = current speed ÷ rated speed × 100%, for example, when the current speed is 600 r / min and the rated speed is 1500 r / min, the load rate = 600 ÷ 1500 × 100% = 40%, and a value lower than this is considered low load); the "design compression ratio" is a preset rated value (e.g., 3.0) based on structural parameters such as the number of teeth and meshing clearance of the first-stage compressor rotor; the "target compression ratio = design compression ratio × 0.8" means that by reducing the target pressure of the first-stage compression, the mechanical energy input during rotor meshing is reduced, the amount of gas internal energy conversion is reduced, thereby suppressing the rise in the first-stage exhaust temperature. 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 of 0.6 MPa to 0.48 MPa, the first-stage compressor load is reduced by 20%, and the temperature rise trend is slowed down. "The reference opening of the first pneumatic regulating valve is increased to 60% to 80%." This range is determined based on the cold air recirculation requirement under low load. When the opening is less than 60%, the cold air recirculation flow is insufficient and cannot effectively dilute the intake temperature of the first-stage intake port. When it is greater than 80%, the excessive recirculation flow causes the first-stage intake pressure to be lower than 0.08MPa, affecting the first-stage compression. For example, taking a reference opening of 70% is 20% higher than 50% under medium load conditions. The cold air recirculation flow increases by 20%, the intake temperature of the first-stage intake port drops from 35℃ to 30℃, and the first-stage exhaust temperature can be controlled within 115℃ to avoid overheating.

[0119] The function of the medium load adjustment unit is to adapt to working conditions with load rates between 40% and 70%. The core is to maintain the stability of interstage parameters by "target compression ratio equal to design compression ratio + PID closed-loop control". The physical meaning of "target compression ratio equals design compression ratio" is: under medium load, the first-stage compression efficiency and the second-stage intake demand are balanced. Maintaining the design compression ratio (e.g., 3.0) ensures that the total exhaust volume reaches the design value (e.g., 10 m³ / min), while avoiding excessive first-stage load or insufficient second-stage intake. "Activate proportional-integral-derivative closed-loop control". The core of this control logic is to dynamically output the opening adjustment amount by calculating the deviation of the interstage compression ratio correction amount in real time. The "proportional coefficient" and "integral coefficient" are preset parameters based on the characteristics of medium load conditions (the proportional coefficient ranges from 0.5 to 0.8, and the integral coefficient ranges from 0.1 to 0.3, such as a proportional coefficient of 0.6 and an integral coefficient of 0.2). "Opening adjustment amount = proportional coefficient × interstage compression ratio correction amount + integral coefficient × ∫ interstage compression ratio correction amount dt" (the integral interval is the first 10 seconds before the current moment to ensure coverage of short-term fluctuations). For example, if the interstage compression ratio correction is 2.0 and the integral value is 5.0, then the opening adjustment = 0.6 × 2.0 + 0.2 × 5.0 = 1.2 + 1.0 = 2.2%, meaning the first pneumatic control valve increases by 2.2% from the base opening of 50%, adjusting to 52.2%. The proportional term quickly responds to the current correction, and the integral term eliminates long-term accumulated deviations, avoiding the deviation superposition caused by traditional open-loop control. This reduces the first-stage exhaust temperature fluctuation from ±8℃ to ±3℃ and the first-stage exhaust pressure fluctuation from ±0.05MPa to ±0.02MPa, significantly improving parameter stability under medium-load conditions.

[0120] The high-load regulating unit is designed to adapt to operating conditions with a load rate greater than 70%. Its core function is to meet the high intake requirements of the second stage by increasing the target compression ratio and the reference opening of the second pneumatic regulating valve. The formula "target compression ratio = design compression ratio × 1.2" means that under high load, the intake volume requirement of the second stage compressor head increases by 30%-40%. By increasing the target compression ratio, the first-stage exhaust pressure can be increased, providing sufficient intake pressure for the second stage compressor head. For example, with a design compression ratio of 3.0, the target compression ratio = 3.0 × 1.2 = 3.6. The first-stage exhaust pressure increases from 0.6 MPa to 0.72 MPa, and the second-stage compressor head intake pressure meets the design requirements (0.6-0.8 MPa). The overall compressor exhaust pressure remains at the design value of 0.8 MPa. "The reference opening of the second pneumatic regulating valve is increased to 70% to 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 and cannot stabilize the secondary intake pressure. When it is greater than 90%, the excessive air supply causes the primary exhaust pressure to exceed the maximum safe pressure of 0.52MPa. For example, taking a reference opening of 80% is 40% higher than the 40% under medium load conditions. The secondary exhaust air supply increases by 40%, the primary exhaust pressure stabilizes at 0.72MPa, and the secondary compressor head compression reaches 110% of the design value, meeting the downstream high air demand.

[0121] The function of the mode switching unit is to achieve a smooth transition when switching between different load modes. The core is to avoid sudden changes in parameters by limiting the opening change rate to no more than 5% per second. The trigger condition for "mode switching" is when the load rate crosses the interval threshold (e.g., when it rises from 38% to 42%, it triggers a switch from low load to medium load; when it drops from 72% to 68%, it triggers a switch from high load to medium load). The "opening change rate" is calculated by "(target opening - current opening) ÷ transition time", and the transition time must meet the requirement that the change rate is ≤5% / second. For example, when switching from low load to medium load, if the first pneumatic control valve is currently open at 70% and the target opening is 50%, with a difference of -20%, then the transition time is 20% ÷ 5% / second = 4 seconds, meaning the opening decreases from 70% to 50% at a rate of 5% per second. When switching from high load to medium load, if the second pneumatic control valve is currently open at 80% and the target opening is 40%, with a difference of -40%, then the transition time is 40% ÷ 5% / second = 8 seconds, with the opening decreasing at a rate of 5% per second. By controlling the rate of change of the opening, the gas flow and pressure in the pipeline change smoothly, avoiding pressure shocks caused by instantaneous flow fluctuations (e.g., a pressure shock of 0.1 MPa during a traditional sudden change, which decreases to 0.02 MPa after a gradual change). Simultaneously, it prevents temperature overshoot due to sudden flow changes (e.g., a temperature overshoot of 5°C during a traditional sudden change, which disappears after a gradual change), ensuring that interstage parameters remain within a safe range during mode switching.

[0122] Through the synergistic effect of the above units, precise adjustment across the entire load range is achieved: 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 temperature is controlled at 115℃, which is lower than the traditional solution (130℃); In the medium load range (load rate 50%), PID closed-loop control reduces the first-stage exhaust temperature fluctuation to ±3℃, which is 62.5% lower than the traditional open-loop control (±8℃); 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 pressure is stabilized at 0.8MPa. During mode switching, the opening change rate is controlled at 5% / second, the pressure fluctuation amplitude is ≤0.02MPa, and there is no temperature overshoot, which is a significant optimization compared to the traditional sudden change solution (pressure fluctuation 0.1MPa, temperature overshoot 5℃). Overall, through load partitioning adaptation, closed-loop fine-tuning, and smooth switching, the system's adjustment capability across the entire load range has been improved, providing key support for the efficient and stable operation of the unit under all operating conditions.

[0123] In one embodiment of the present invention, the temperature and pressure monitoring module further includes:

[0124] The third calculation unit is used to calculate the acceleration of the change in the first-stage exhaust temperature. The acceleration is calculated by the change in the rate of temperature change per unit time, and the unit is degrees Celsius per second squared.

[0125] The generation unit is used to generate an advance adjustment signal when the current changing acceleration is greater than the preset maximum value of the changing acceleration (the preset maximum value of the changing acceleration is 0.5 degrees Celsius per second squared). The signal value is calculated by multiplying the acceleration by the current opening degree and then by a coefficient of 0.05 to 0.15.

[0126] The superposition unit is used to superimpose the advance adjustment signal onto the opening command of the first pneumatic control valve; the new opening is equal to the current opening plus the advance adjustment signal.

[0127] The protection unit is used to activate the over-adjustment protection and freeze the opening of the first pneumatic regulating valve for 10 to 15 seconds when the current variable acceleration is less than the preset minimum value of variable acceleration (the preset minimum value of variable acceleration is -0.3 degrees Celsius per second squared).

[0128] As described above, by constructing a predictive temperature control logic, the third calculation unit calculates the temperature change acceleration based on the difference in the rate of temperature change per unit time. The generation unit generates an advance adjustment signal when the acceleration exceeds the preset maximum value. The superposition unit integrates the signal into the opening command of the first pneumatic regulating valve to intervene in the temperature change trend in advance. The protection unit freezes the opening when the acceleration is lower than the preset minimum value to avoid over-adjustment. This achieves accurate capture and predictive control of the rate of temperature change in the first stage exhaust, solves the adjustment lag and over-adjustment risk problems that still exist based on temperature change rate compensation, further improves the system's response time and adjustment stability to dynamic temperature changes, and ensures that the interstage temperature is always stable within a safe and efficient range.

[0129] The core function of the third calculation unit is to calculate the acceleration of the change in the first-stage exhaust temperature in real time. This parameter is calculated by "acceleration = (current temperature change rate - previous temperature change rate) ÷ time interval". The "current temperature change rate" and "previous temperature change rate" are both obtained from the monitoring results of the detection unit. The "time interval" is the time difference between two rate change acquisitions (fixed at 1 second, which can accurately reflect the dynamic change of the rate of change and avoid data noise interference caused by too short an interval). The unit of the calculation result is degrees Celsius per second squared. Acceleration directly quantifies the rate of temperature change. For example, if the current rate of temperature change is 2.6℃ / s and the rate of change was 2.0℃ / s one second ago, then the acceleration is (2.6 - 2.0) ÷ 1 = 0.6℃ / s², indicating that the rate of temperature increase is accelerating. If the current rate of change is 1.4℃ / s and the rate of change was 2.0℃ / s one second ago, then the acceleration is (1.4 - 2.0) ÷ 1 = -0.6℃ / s², indicating that the rate of temperature increase is rapidly slowing down. The acceleration data calculated by this unit is transmitted to the generation and protection units in real time. Its significance lies in upgrading the dynamic trend of temperature change from a first-order rate of change to a second-order acceleration, providing a core basis for subsequent anti-adjustment and over-adjustment protection to reflect the rate of trend, avoiding the lagging judgment of traditional methods that rely solely on the rate of change.

[0130] The function of the generation unit is to generate an advance adjustment signal when the acceleration exceeds a preset maximum value, thereby increasing the opening adjustment force in advance and offsetting the adjustment lag caused by the intensified temperature change trend. The preset maximum value for the acceleration is set to 0.5℃ / s², which is determined based on the temperature regulation inertia characteristics of the oil-free screw air compressor. When the acceleration exceeds 0.5℃ / s², without advance intervention, the temperature change rate will rise to over 2.5℃ / s within 5 seconds, and the first-stage exhaust temperature will exceed the safety threshold. The advance adjustment signal value is calculated as: "Acceleration × Current Opening × Preset Coefficient," where "Current Opening" is the corrected opening of the first pneumatic control valve (e.g., 25%, which incorporates load rate and thermal inertia compensation), and the "Preset Coefficient" ranges from 0.05 to 0.15 (e.g., 0.1, which ensures the signal value effectively intervenes in advance while avoiding over-adjustment). A larger acceleration indicates a more pronounced temperature change trend, requiring a larger advance signal; a larger current opening indicates a stronger current cooling capacity, requiring a corresponding signal amplitude to maintain stable adjustment. For example, if the acceleration change is 0.6℃ / s² (exceeding the preset maximum value of 0.5℃ / s²), the current opening is 25%, and the coefficient is 0.1, then the advance adjustment signal value = 0.6 × 25% × 0.1 = 1.5%. This signal indicates that an additional 1.5% needs to be added on the basis of the current opening to enhance the cold air return flow in advance to cope with the upcoming intensified temperature rise trend.

[0131] The function of the superposition unit is to integrate the advance adjustment signal into the opening command of the first pneumatic control valve to form the final executed opening value. The adjustment logic is "new opening of the first pneumatic control valve = current opening + advance adjustment signal value". Here, "current opening" is the corrected opening (e.g., 25%), and "advance adjustment signal value" is the calculation result of the generation unit (e.g., 1.5%). Therefore, new opening = 25% + 1.5% = 26.5%. By directly superimposing the advance signal onto the current opening, the valve increases its opening in advance (or decreases its opening in advance when the trend is downward), allowing the adjustment of the cold air return flow to precede the intensification of the temperature change trend, thereby offsetting the lag caused by adjustment inertia. For example, without a superimposed signal, the cold air recirculation flow rate corresponding to the current opening degree 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 1.5% signal, the recirculation flow rate of the opening degree 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 superimposed mechanism can reduce the temperature overshoot range from the traditional 8-12℃ to 3-4℃, shorten the regulation lag time, and significantly improve response timeliness.

[0132] The protection unit's function is to address the risk of over-adjustment when the acceleration change is below a preset minimum value. When the acceleration change is less than -0.3℃ / s² (this preset minimum value is determined based on compression efficiency requirements; when the acceleration is below -0.3℃ / s², the temperature drop rate will increase from 1℃ / s to 1.9℃ / s within 3 seconds, easily leading to temperatures below the high-efficiency compression range of 100℃), the over-adjustment protection is activated, freezing the opening of the first pneumatic regulating valve for 10 to 15 seconds (e.g., 12 seconds, this duration was determined experimentally to both suppress excessive temperature drop and avoid new fluctuations caused by prolonged freezing). For example, if the acceleration change is -0.4℃ / s² (below -0.3℃ / s²), and the current opening of the first pneumatic regulating valve is 26.5%, the protection unit will immediately freeze this opening for 12 seconds, during which time the opening will not be adjusted according to temperature changes. When the temperature drops too quickly, the adjustment logic that continuously increases the return flow (or decreases the opening) is cut off by freezing the opening to prevent the temperature from dropping excessively. If not frozen, the opening will continue to increase to 28%, the return flow will increase further, and the temperature will drop from 120°C to 95°C within 12 seconds, and the first-stage compression efficiency will drop from 85% to 70%. After freezing, the temperature will only drop to 105°C, which is still within the high-efficiency range, and the compression efficiency will remain within the high-efficiency range. At the same time, it avoids parameter fluctuations caused by subsequent reverse adjustment due to excessively low temperature.

[0133] Through the synergistic effect of the aforementioned units, predictive control of the first-stage exhaust temperature change trend is achieved: in scenarios where the temperature rise trend intensifies (acceleration 0.6℃ / s²), the advance adjustment signal reduces the temperature overshoot amplitude and shortens the adjustment lag time; in scenarios where the temperature falls too rapidly (acceleration -0.4℃ / s²), the overshoot protection maintains the minimum temperature at 105℃, reducing compression efficiency loss; in scenarios with normal fluctuations (acceleration -0.2 to 0.4℃ / s²), the system does not need to activate the advance or protection mechanism, but only maintains the original compensation logic to ensure smooth adjustment. Overall, by monitoring and predictively controlling the acceleration of temperature changes, the technology fills the gap in traditional methods that rely solely on rate-of-change adjustment, further improving the system's control accuracy and stability under dynamic temperature change scenarios, and providing more refined temperature protection for the efficient and safe operation of the unit.

[0134] In one embodiment of the present invention, a calculation module is further included, the calculation module comprising:

[0135] The fourth calculation unit is used to obtain the first-stage intake temperature and first-stage intake pressure through the intake temperature sensor and the intake pressure sensor, and to obtain the first-stage exhaust temperature and first-stage exhaust pressure through the exhaust temperature sensor and the exhaust pressure sensor, and to calculate the actual compression ratio based on the first-stage intake temperature, first-stage exhaust temperature, first-stage intake pressure and first-stage exhaust pressure.

[0136] The specific calculation process is as follows: To calculate the actual compression ratio based on the ideal gas equation, firstly, the first-stage intake temperature, first-stage exhaust temperature, first-stage intake pressure, and first-stage exhaust pressure are obtained. Then, the exhaust pressure is divided by the intake pressure and multiplied by (intake temperature divided by exhaust temperature) raised to the power of (adiabatic index divided by adiabatic index minus 1). The adiabatic index is between 1.30 and 1.40.

[0137] The fifth calculation unit is used to calculate the compression efficiency based on the actual compression ratio.

[0138] Compression efficiency is calculated by dividing the actual compression ratio by the power of (adiabatic index minus 1 divided by the adiabatic index) minus 1 by the power of (adiabatic index minus 1 divided by the adiabatic index) minus 1, and then multiplying by 100%.

[0139] The sixth calculation unit is used to trigger the efficiency optimization mode when the compression efficiency is less than the preset efficiency value (the preset efficiency value is 75%). The specific optimization method is to reduce the set temperature value by 10 to 20 degrees Celsius and reduce the designed compression ratio by 0.2 to 0.5.

[0140] As described above, a closed-loop efficiency control system is constructed. The fourth calculation unit calculates the actual compression ratio based on the temperature and pressure parameters of the first-stage intake and exhaust. The fifth calculation unit quantifies the compression efficiency by comparing the actual compression ratio with the design compression ratio. When the compression efficiency is lower than the preset value, the sixth calculation unit triggers the efficiency optimization mode and adjusts the set temperature value and the design compression ratio. This enables real-time monitoring and dynamic optimization of the compression efficiency of the oil-free screw air compressor. This solves the problem of focusing only on interstage pressure and temperature control and not actively controlling compression efficiency in the above content. It ensures that the unit maintains stable interstage parameters while always operating in the high-efficiency range, reducing ineffective energy consumption and improving operating economy.

[0141] The core function of the fourth calculation unit is to calculate the actual compression ratio based on the temperature and pressure parameters of the first-stage intake and exhaust. This calculation process is based on the ideal gas law and fully considers the coupling relationship between temperature and pressure during gas compression. Specifically, the required "first-stage intake temperature" and "first-stage intake pressure" are obtained from the intake temperature sensor (accuracy ±0.5℃) and intake pressure sensor (accuracy ±0.01MPa) installed at the first-stage intake port in the temperature and pressure monitoring module. The "first-stage exhaust temperature" and "first-stage exhaust pressure" are obtained from the exhaust temperature sensor and exhaust pressure sensor installed at the first-stage exhaust port. All parameters are collected at a frequency of 1 time / second to ensure that the data reflects the gas compression state in real time. The specific calculation formula is as follows:

[0142] ;

[0143] in, Indicates the actual compression ratio. Indicates the first-stage exhaust pressure. Indicates the first-stage inspiratory pressure. Indicates the first-stage inhalation temperature. Indicates the first-stage exhaust temperature. Indicates the adiabatic index;

[0144] The "adiabatic index" ranges from 1.30 to 1.40 (for air, this value is determined experimentally, with 1.35 being a typical value; it reflects the heat transfer characteristics of air during compression. The higher the adiabatic index, the more significant the temperature rise during gas compression). During gas compression, pressure and temperature changes are coupled. The pressure ratio alone cannot accurately reflect the actual degree of compression; therefore, a temperature ratio and the adiabatic index must be introduced for correction to obtain a compression ratio that closely reflects the actual compression process. For example, if the first-stage intake temperature is 300K (27℃) and the first-stage intake pressure is 0.1MPa, the first-stage exhaust temperature is 360K (87℃) and 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. This value indicates that the current actual compression ratio is lower than the design value of 3.0, and there is a risk of efficiency reduction.

[0145] The function of the fifth calculation unit is to quantitatively calculate the current compression efficiency by comparing the actual compression ratio with the design compression ratio, establishing 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%", where "Design compression ratio" is the rated compression ratio determined during the unit design phase (e.g., 3.0, stored in the PLC control module parameter library), and "Adiabatic index" is the same as the value taken by the fourth calculation unit (e.g., 1.35). 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 (at the design compression ratio). 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, if 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%. This value 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.

[0146] 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 restore efficiency by adjusting the core control parameters. The "preset efficiency value of 75%" is determined based on the energy consumption economy and equipment safety boundary of the oil-free screw air compressor. When the efficiency is lower than 75%, the energy consumption per unit displacement will exceed the industry's energy efficiency level 2 standard (GB19153-2019), and long-term operation will aggravate the wear of the compressor head. The 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". The two work together: reducing the set temperature value (e.g., from 120°C to 105°C) can enhance the cold airflow backflow regulation space of the first pneumatic regulating valve, suppress the temperature rise caused by the decrease in efficiency, and avoid further deterioration of leakage due to the temperature rise; reducing the design compression ratio (e.g., from 3.0 to 2.7) can reduce the compression load of the first-stage compressor head, making the actual compression ratio closer to the adjusted design compression ratio, and reducing energy waste caused by the deviation of the compression ratio. For example, when the compression efficiency drops to 65.62%, the optimization mode is triggered: the set temperature value drops from 120℃ to 105℃, and the design compression ratio drops from 3.0 to 2.7. At this time, the fourth calculation unit recalculates the actual compression ratio and increases it to 2.5 (because the design compression ratio is reduced, the deviation between the actual and the design is reduced). 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 significantly recovers to the high-efficiency range, the energy consumption per unit exhaust volume decreases by 16%, and the first-stage exhaust temperature drops from 135℃ to 110℃, avoiding the rotor from seizing due to excessive temperature.

[0147] Through the synergistic effect of the aforementioned units, full-cycle control of compression efficiency is achieved: During the efficiency monitoring phase, the fourth and fifth calculation units can quantify the actual compression ratio and efficiency in real time, with a monitoring accuracy of ±2%, achieving visualized efficiency management compared to traditional no-monitoring schemes; During the efficiency optimization phase, parameter adjustments by the sixth calculation unit can improve efficiency and reduce energy consumption without requiring downtime, avoiding production interruptions caused by maintenance; During long-term operation, this module can dynamically adapt to the decrease in actual compression ratio caused by die head wear, extending the unit's efficient operating cycle (from the traditional 2000 hours to 3500 hours), reducing maintenance frequency and costs. Overall, through efficiency quantification and dynamic optimization, the gap in efficiency management of traditional control schemes is filled, achieving the dual goals of stable inter-stage parameters and high compression efficiency, providing key technical support for the economical operation of the unit.

[0148] In one embodiment of the present invention, a computer device is also disclosed, including a memory and a processor. The memory stores a computer program, and the processor executes the computer program to realize the operation of an interstage pressure and temperature flexible control system for an oil-free screw air compressor.

[0149] In one embodiment of the present invention, a computer-readable storage medium is also disclosed, on which a computer program is stored, which, when executed by a processor, enables the operation of an interstage pressure and temperature flexible control system for an oil-free screw air compressor.

[0150] Those skilled in the art will understand that all or part of the processes in the above embodiments can be implemented by a computer program instructing related hardware. The computer program can be stored in a non-volatile computer-readable storage medium. When executed, the computer program can include the processes of the embodiments of the above methods. Any references to memory, storage, databases, or other media used in this application and 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. By way of illustration and not limitation, RAM is available in a variety of forms, such as static RAM (SRAM), dynamic RAM (DRAM), synchronous DRAM (SDRAM), dual-speed SDRAM (SSRSDRAM), enhanced SDRAM (ESDRAM), synchronous link DRAM (SLDRAM), RAMbus direct RAM (RDRAM), direct memory bus dynamic RAM (DRDRAM), and memory bus dynamic RAM (RDRAM).

[0151] It should be noted that, in this document, the terms "comprising," "including," or any other variations thereof are intended to cover non-exclusive inclusion, such that a process, apparatus, article, or method that comprises a list of elements includes not only those elements but also other elements not expressly listed, or elements inherent to such process, apparatus, article, or method. Unless otherwise specified, an element defined by the phrase "comprising one..." does not exclude the presence of other identical elements in the process, apparatus, article, or method that includes that element.

[0152] The above description is merely a preferred embodiment of the present invention and does not limit the scope of this application. Any equivalent results or equivalent process transformations made based on the content of the present invention's specification and drawings, or direct or indirect applications in other related technical fields, are similarly included within the scope of protection of this application.

Claims

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

2. The interstage pressure and temperature flexible control system for the oil-free screw air compressor according to claim 1, characterized in that, The PLC control module includes: The data acquisition unit is used to acquire the load rate of the air compressor in real time. The load rate of the air compressor is calculated by comparing the current speed with the rated speed. The first calculation unit is used to calculate the load correction factor, which is determined based on the load rate segmentation. The first correction unit is used to correct the opening degree of the first pneumatic control valve by multiplying the initial opening degree by the load correction coefficient to obtain the corrected opening degree of the first pneumatic control valve. The second correction unit is used to correct the opening degree of the second pneumatic control valve. The corrected opening degree of the second pneumatic control valve is obtained by multiplying the initial opening degree by 2 and subtracting the load correction coefficient. The forced correction unit is used to forcibly increase the opening of the corrected first pneumatic regulating valve by 20% to 35% when the actual exhaust temperature exceeds the alarm threshold.

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

4. The interstage pressure and temperature flexible control system for 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, wherein the minimum safety pressure is 0.7 times the design pressure and the maximum safety pressure is 1.3 times the design pressure; The first opening adjustment unit is used to adjust the opening of the first pneumatic regulating valve when the actual first-stage exhaust pressure is less than the minimum safe pressure. The second opening adjustment unit is used to adjust the opening of the second pneumatic regulating valve when the actual first-stage exhaust pressure is greater than the maximum safe pressure. The third opening adjustment unit is used to close the second pneumatic regulating valve and fully open the first pneumatic regulating valve when the actual exhaust temperature exceeds 180 degrees Celsius.

5. The interstage pressure and temperature flexible control system for the oil-free screw air compressor according to claim 1, characterized in that, The graded adjustment module includes: The low-load regulating unit is used to set the target compression ratio to 0.8 times the design compression ratio in low-load mode and increase the reference opening of the first pneumatic regulating valve to the corresponding range. The medium load adjustment unit is used to set the target compression ratio to equal the design compression ratio in medium load mode and enable closed-loop control. The opening adjustment amount is calculated through the closed-loop control. It is calculated by multiplying the proportional coefficient by the interstage compression ratio correction and then integrating the integral coefficient by the interstage compression ratio correction. The high-load regulating 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 regulating valve to the corresponding range. The mode switching unit is used to employ a gradual transition method during mode switching, with the opening change rate limited to an increase or decrease of no more than 5% per second.

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

7. The interstage pressure and temperature flexible control system for the oil-free screw air compressor according to claim 1, characterized in that, It also includes a computing module, which includes: The fourth calculation unit is used to obtain the first-stage intake temperature and first-stage intake pressure through the intake temperature sensor and the intake pressure sensor, and to obtain the first-stage exhaust temperature and first-stage exhaust pressure through the exhaust temperature sensor and the exhaust pressure sensor, and to calculate the actual compression ratio based on the first-stage intake temperature, first-stage exhaust temperature, first-stage intake pressure and first-stage exhaust pressure. The fifth calculation unit is used to calculate the compression efficiency based on the actual compression ratio. The sixth calculation unit is used to trigger the efficiency optimization mode when the compression efficiency is less than the preset efficiency value.

8. A computer device comprising a memory and a processor, wherein the memory stores a computer program, characterized in that, When the processor executes the computer program, it enables the operation of the system according to any one of claims 1 to 7.

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

Images