Control method of multi-stage liquid-driven piston compressor system

Abstract

The invention discloses a control method of a multi-stage liquid-driven piston compressor system. The control method comprises the following steps: step 1, calculating geometric parameters of each stage of compressor; secondly, the pressure of the air inlet end is recognized through an acquisition card, and the pressure ratio of each stage is calculated according to the required exhaust pressure value, the compressor stage number and other parameters; thirdly, the rotating speed ratio of each stage of compressor is calculated according to the pressure ratio, the volume of each stage of cylinder and the number of cylinders; fourthly, the maximum compressor rotating speed is screened out according to the rotating speed ratio of all stages of compressors with the maximum motor power and the maximum pump station oil pressure parameter as limits; and fifthly, the reversing frequency value of a reversing mechanism is calculated according to the rotating speed of each stage of compressor, and therefore the effect of changing design working conditions is achieved.

Description

Technical Field

[0001] The present invention relates to the technical field of compressors, and more particularly to a control method for a multi-stage liquid-driven piston compressor system. Background Art

[0002] Hydrogen is a secondary energy source with a wide range of sources, is clean and low-carbon, flexible and efficient, and has a wide range of applications. Hydrogen is a key medium for China's energy structure transformation. As professional equipment for filling hydrogen fuel cells, hydrogen refueling stations currently have two filling standards: 35MPa and 70MPa, depending on the hydrogen supply pressure level. Operating in a high-pressure environment makes components extremely susceptible to damage. Due to the wide explosion range of hydrogen, the compression process must be pollution-free, further limiting the development of hydrogen compressors. Therefore, the development of highly reliable and pollution-free hydrogen compressors has become the key to promoting the development of the hydrogen energy industry. Liquid-driven piston compressors have the advantages of load start and stop, high volumetric efficiency, few wearing parts, and simple maintenance. They will be one of the main applications of hydrogen compressors in future hydrogen refueling stations.

[0003] As the core power component of a hydrogen refueling station, the compressor's operating efficiency determines the efficiency of the entire system. This efficiency is primarily determined by both design and operating conditions. When operating conditions match design conditions, the unit operates at high efficiency. Because hydrogen at a hydrogen refueling station is sourced from various sources, including industrial exhaust gas and natural gas, the hydrogen pressure received by the station varies. The compressor often needs to operate under non-design conditions, resulting in lower compressor efficiency and a lower overall pressurization system efficiency, leading to poor economic performance at the hydrogen refueling station.

[0004] Adjusting operating conditions involves adjusting the geometry and control methods. However, adjusting the geometry is costly and can pose safety risks, necessitating an urgent need for a control system specifically tailored to the liquid-driven compressors in hydrogen refueling stations. The lack of a mature control solution and control system in China has led to long-term low reliability and poor economic performance for hydrogen refueling stations, severely hindering the industrialization of hydrogen energy in my country.

[0005] Due to the diverse gas transmission sources, intake pressure often differs from the designed operating conditions. Most liquid-driven piston compressors in hydrogen refueling stations operate under a specific operating condition (speed) as designed. These compressors often operate outside of their designed operating conditions, resulting in excessively high compressor exhaust temperatures, high pumping station oil pressures, and significantly reduced component lifespans. This leads to low unit reliability and poor economic efficiency. Adjustments for varying intake pressures are impossible. Summary of the Invention

[0006] The object of the present invention is to provide a control method for a multi-stage liquid-driven piston compressor system, in order to solve the technical problems existing in the background technology.

[0007] In order to achieve the above object, the present invention adopts the following technical solutions:

[0008] A control method for a multi-stage liquid-driven piston compressor system, comprising:

[0009] The first step is to calculate the geometric parameters of each stage of the compressor;

[0010] The second step is to identify the pressure at the inlet end through the acquisition card, and calculate the pressure ratio of each stage based on the required exhaust pressure value and parameters such as the number of compressor stages;

[0011] The third step is to calculate the speed ratio of each compressor according to the pressure ratio, cylinder volume of each level, and number of cylinders;

[0012] The fourth step is to select the maximum compressor speed based on the speed ratio of each compressor, with the maximum motor power and the maximum oil pressure of the pump station as the limit;

[0013] The fifth step is to calculate the commutation frequency value of the commutation mechanism according to the speed of each level of compressor, so as to achieve the effect of changing the design working conditions.

[0014] In some embodiments, the first step of calculating geometric parameters of each stage of the compressor includes:

[0015] V=π*d 2 *1

[0016] Where: V is the cylinder volume, d is the cylinder radius, and l is the cylinder stroke.

[0017] In some embodiments, the second step is to identify the pressure at the air inlet end through the acquisition card and calculate the pressure ratio of each stage according to the required exhaust pressure value and parameters such as the number of compressor stages; including:

[0018]

[0019] Where: ε is the pressure ratio, p 1s is the pressure at the first stage inlet, p 2d is the pressure at the secondary exhaust end.

[0020] In some embodiments, the third step of calculating the speed ratio of each compressor stage based on the pressure ratio, the cylinder volume of each stage, and the number of cylinders comprises:

[0021]

[0022] Where: n is the compression speed, V is the compressor cylinder volume, g is the number of compressor cylinders, ε is the pressure ratio, the superscript 1 represents the first stage, and the superscript 2 represents the second stage.

[0023] In some embodiments, the fourth step, based on the maximum motor power and the maximum oil pressure parameters of the pump station, selects the maximum compressor speed according to the speed ratio of each compressor; including:

[0024]

[0025] Where: P is the motor power, V is the compressor cylinder volume, g is the number of compressor cylinders, n is the compression speed, p is the intake pressure, k is the process index, ε is the pressure ratio, η ad is the adiabatic efficiency, η d for mechanical efficiency;

[0026]

[0027] Where: p o is the working oil pressure of the pump station, p s is the pressure at the inlet end, p d is the pressure at the exhaust end, S g is the end surface area of ​​the compressor cylinder, S o is the end surface area of ​​the compressor cylinder, p omax The maximum working pressure of the oil pump is determined according to the oil pump model, generally 25-40Mpa, p omin It is the minimum working pressure of the oil pump, generally 2-3Mpa, to prevent the pipeline from being sucked into negative pressure by the oil pump, leaking gas, and causing cavitation of the oil pump.

[0028] After calculating the compressor speed, due to the compressibility of the hydraulic oil, the actual displacement of the hydraulic oil is less than the theoretical displacement. It is necessary to consider the impact of the volume reduction and correct the speed.

[0029] V0=K×V0×(p o -p omin )

[0030]

[0031] Where: V′0 is p o The hydraulic volume of the pressure, K is the bulk modulus, and V0 is p omin The volume of hydraulic oil under pressure, n' is the corrected speed.

[0032] Compared with the prior art, the present invention has the following beneficial effects:

[0033] (1) The control method of the multi-stage liquid-driven piston compressor system designed in the present invention can change the design operating conditions from a specific point to a range, thereby widening the operating condition range of the liquid-driven piston compressor, allowing the unit to operate efficiently under different operating conditions, and improving the working efficiency of the system.

[0034] (2) By adjusting the pressure ratio, the maximum oil pressure required under the same conditions is reduced, thus reducing the cost of the oil pump; the exhaust pressure is reduced, which increases the service life of the seals. Overall, the construction and operation costs of the unit are reduced. BRIEF DESCRIPTION OF THE DRAWINGS

[0035] Figure 1 Multi-stage liquid-driven piston compressor flow chart;

[0036] Figure 2 Control logic diagram of a multi-stage liquid-driven piston compressor. DETAILED DESCRIPTION

[0037] In order to make the purpose, technical solutions and advantages of the present application clearer, the technical solutions in the embodiments of the present application will be described in more detail below in conjunction with the drawings in the preferred embodiments of the present application. In the drawings, the same or similar reference numerals throughout represent the same or similar parts or parts with the same or similar functions. The described embodiments are part of the embodiments of the present application, not all of the embodiments. The embodiments described below with reference to the drawings are exemplary and are intended to be used to explain the present application, and should not be understood as limitations on the present application. Based on the embodiments in the present application, all other embodiments obtained by ordinary technicians in this field without making creative work are within the scope of protection of this application.

[0038] The embodiments of the present application are described in detail below with reference to the accompanying drawings.

[0039] In the description of this application, it should be noted that, unless otherwise expressly specified or limited, the terms "mounted," "connected," and "connected" should be understood in a broad sense. For example, they may refer to a fixed connection, an indirect connection via an intermediate medium, internal communication between two components, or an interaction between two components. Those skilled in the art will understand the specific meanings of the above terms in this application based on the specific circumstances.

[0040] In the description of this application, it should be understood that the terms "upper", "lower", "front", "back", "vertical", "horizontal", "top", "bottom", "inside", "outside", etc., indicating orientations or positional relationships, are orientations or positional relationships based on the accompanying drawings, and are only for the convenience of describing this application and simplifying the description, rather than indicating or implying that the device or element referred to must have a specific orientation, be constructed and operated in a specific orientation, and therefore should not be understood as a limitation on this application.

[0041] In addition, the terms "comprises" and "having" and any variations thereof are intended to cover a non-exclusive inclusion; for example, a process, method, system, product, or display that comprises a series of steps or elements is not necessarily limited to those steps or elements expressly listed but may include other steps or elements not expressly listed or inherent to such process, method, product, or display.

[0042] The following will be combined Figure 1-Figure 2, a control method for a multi-stage liquid-driven piston compressor system involved in an embodiment of the present application is described in detail. It is worth noting that the following embodiments are only used to explain the present application and do not constitute a limitation of the present application.

[0043] Embodiment 1:

[0044] like Figure 1-2 As shown, a control method for a multi-stage liquid-driven piston compressor system is used to control the multi-stage liquid-driven piston compressor system. The flow chart of the multi-stage liquid-driven piston compressor system is shown in FIG. Figure 1 As shown, it includes a motor 1, a transmission mechanism 2, a two-way oil pump 3, an oil tank 4, an oil overflow valve 5, a one-way valve 6, an oil replenishment pump 7, a first-stage compressor reversing valve 8, a second-stage compressor reversing valve 9, a pressure sensor 10, a first-stage compressor 11, a second-stage compressor 12, and an automatic control system 13.

[0045] The main shaft of motor 1 is connected to the main shafts of bidirectional oil pump 3 and charge pump 7 via transmission mechanism 2. Bidirectional oil pump 3 provides the main oil supply to the oil circuit, while charge pump 7 prevents low hydraulic oil pressure in the oil circuit, which could cause the pipeline to be compressed by air. Because both are heavily loaded and need to withstand significant torsional forces, transmission mechanism 2 is constructed of solid steel. On the supply side of bidirectional oil pump 3, the pipeline is subject to a high pressure of 30 MPa. Ultra-high-pressure hydraulic hoses are used between bidirectional oil pump 3, the oil relief valve 5, the check valve 6, the first-stage compressor reversing valve 8, and the first-stage compressor 11 to mitigate pipeline resistance during oil supply and reduce oil pressure pulsation. On the return side of charge pump 7, the internal pressure in the pipeline is generally 2-3 MPa. Conventional thin-walled steel pipes are used to rigidly connect the charge pump 7, the oil relief valve 5, and the check valve 6. The pipes are kept short to prevent pipeline resonance from affecting pump operation. The same applies to the second-stage compressor.

[0046] The automatic control system 13 is connected to the pressure sensor 10 using a two-core signal line to obtain relevant pressure parameters. The automatic control system 13 is connected to the first-stage compressor reversing valve 8 and the second-stage compressor reversing valve 9 using a four-core signal line, with two cores providing power supply and two cores inputting control signals.

[0047] The control method of the multi-stage liquid-driven piston compressor system designed in the present application uses an acquisition card to identify the signal of the pressure sensor at the air inlet, transmits the signal to the control system, uses the program in the control system to calculate the reasonable rotation speed of the first and second stage compressors, and adjusts the mechanism that controls the pump speed to achieve pressure ratio adjustment. This improves the working conditions of the high pressure ratio of the compressor piston, reduces the cylinder exhaust temperature, and improves the efficiency of the unit; reduces the maximum oil pressure of the oil cylinder, and increases the service life of the pump station. Since the cost of the pump station is proportional to the maximum oil pressure, the reduction in oil pressure also reduces the cost of the pump station. Through this control system, the operating range of the liquid-driven piston compressor is widened, the cost of the liquid-driven compressor system is reduced, and the efficient operation of the liquid-driven compressor is achieved.

[0048] The control logic diagram of the multi-stage liquid-driven piston compressor of the present invention is as follows: Figure 2 shown.

[0049] The first step is to calculate the geometric parameters of each stage of the compressor (exhaust pressure p 2d , cylinder capacity V1, V2, etc.).

[0050] V=π*d 2 *1

[0051] Where: V is the cylinder volume, d is the cylinder radius, and l is the cylinder stroke.

[0052] The second step is to identify the pressure at the inlet end through the acquisition card (p 1s ), calculate the pressure ratio (ε) of each stage based on the required exhaust pressure value and the number of compressor stages and other parameters.

[0053]

[0054] Where: ε is the pressure ratio, p 1s is the pressure at the first stage inlet, p 2d is the pressure at the secondary exhaust end.

[0055] The third step is to calculate the speed ratio of each compressor stage based on parameters such as the compression ratio, cylinder volume at each stage, and the number of cylinders (g1, g2).

[0056]

[0057] Where: n is the compression speed, V is the compressor cylinder volume, g is the number of compressor cylinders, ε is the pressure ratio, the superscript 1 represents the first stage, and the superscript 2 represents the second stage.

[0058] The fourth step is to select the maximum compressor speed based on the speed ratio of each compressor, with the maximum motor power and the maximum oil pressure of the pump station as the limit; including:

[0059]

[0060] Where: P is the motor power, V is the compressor cylinder volume, g is the number of compressor cylinders, n is the compression speed, p is the intake pressure, k is the process index, ε is the pressure ratio, η ad is the adiabatic efficiency, η d for mechanical efficiency;

[0061]

[0062] Where: p o is the working oil pressure of the pump station, p s is the pressure at the inlet end, p d is the pressure at the exhaust end, S g is the end surface area of ​​the compressor cylinder, S o is the end surface area of ​​the compressor cylinder, p omax The maximum working pressure of the oil pump is determined according to the oil pump model, generally 25-40Mpa, p omin It is the minimum working pressure of the oil pump, generally 2-3Mpa, to prevent the pipeline from being sucked into negative pressure by the oil pump, leaking gas, and causing cavitation of the oil pump.

[0063] After calculating the compressor speed, due to the compressibility of the hydraulic oil, the actual displacement of the hydraulic oil is less than the theoretical displacement. It is necessary to consider the impact of the volume reduction and correct the speed.

[0064] V'0=K×V0×(p o -p omin )

[0065]

[0066] Where: V′0 is p o The hydraulic volume of the pressure, K is the bulk modulus, and V0 is p omin The volume of hydraulic oil under pressure, n' is the corrected speed.

[0067] The fifth step is to calculate the commutation frequency value of the commutation mechanism according to the speed of each level of compressor, so as to achieve the effect of changing the design working conditions.

[0068] The above description is only a preferred embodiment of the present invention and is not intended to limit the present invention. Any modifications, equivalent substitutions and improvements made within the spirit and principles of the present invention should be included in the scope of protection of the present invention.

Claims

1. A control method for a multi-stage liquid-driven piston compressor system, characterized in that: include: The first step is to calculate the geometric parameters of each stage of the compressor; The second step is to identify the pressure at the inlet end through the acquisition card, and calculate the pressure ratio of each stage based on the required exhaust pressure value and parameters such as the number of compressor stages; The third step is to calculate the speed ratio of each compressor according to the pressure ratio, cylinder volume of each level, and number of cylinders; The fourth step is to select the maximum compressor speed based on the speed ratio of each compressor, with the maximum motor power and the maximum oil pressure of the pump station as the limit; The fifth step is to calculate the commutation frequency value of the commutation mechanism according to the speed of each level of compressor, so as to achieve the effect of changing the design working conditions.

2. The control method of a multi-stage liquid-driven piston compressor system according to claim 1, characterized in that: The first step is to calculate the geometric parameters of each stage of the compressor; including: V=π*d 2 *l; Where: V is the cylinder volume, d is the cylinder radius, and l is the cylinder stroke.

3. The control method of a multi-stage liquid-driven piston compressor system according to claim 1, characterized in that: The second step is to identify the pressure at the inlet end through the acquisition card, and calculate the pressure ratio of each stage based on the required exhaust pressure value and parameters such as the number of compressor stages; include: Where: ε is the pressure ratio, p 1s is the pressure at the first stage inlet, p 2d is the pressure at the secondary exhaust end.

4. The control method of a multi-stage liquid-driven piston compressor system according to claim 1, characterized in that: The third step is to calculate the speed ratio of each compressor according to the pressure ratio, the cylinder volume of each level, and the number of cylinders; including: Where: n is the compression speed, V is the compressor cylinder volume, g is the number of compressor cylinders, ε is the pressure ratio, the superscript 1 represents the first stage, and the superscript 2 represents the second stage.

5. The control method of a multi-stage liquid-driven piston compressor system according to claim 1, characterized in that: The fourth step is to select the maximum compressor speed based on the speed ratio of each level of compressors, with the maximum motor power and the maximum oil pressure parameters of the pump station as restrictions; including: Where: P is the motor power, V is the compressor cylinder volume, g is the number of compressor cylinders, n is the compression speed, p is the intake pressure, k is the process index, ε is the pressure ratio, η ad is the adiabatic efficiency, η d for mechanical efficiency; Where: p o is the working oil pressure of the pump station, p s is the pressure at the inlet end, p d is the pressure at the exhaust end, S g is the end surface area of ​​the compressor cylinder, S o is the end surface area of ​​the compressor cylinder, p omax is the maximum working pressure of the oil pump, p omin It is the minimum working pressure of the oil pump; After calculating the compressor speed, it is necessary to consider the impact of volume reduction and make corrections to the speed; V′0=K×V0×(p o -p omin ); Where: V′0 is p o The hydraulic volume of the pressure, K is the bulk modulus, and V0 is p omin The volume of hydraulic oil under pressure, n' is the corrected speed.

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