Design method of super-high pressure large displacement liquid drive compressor system and compressor system

Abstract

This invention relates to the field of compressor technology, specifically to a design method and system for an ultra-high pressure, high-displacement hydraulically driven compressor system. The invention includes a primary compressor, a secondary compressor, and a tertiary compressor connected in series along the exhaust direction. The parameter design of each compressor includes the following steps: S1, selecting the sealing type of the primary compressor, determining its piston stroke and reciprocating frequency, and determining the intake pressure, maximum exhaust pressure, and displacement of the compressor system; S2, selecting the sealing types of the secondary and tertiary compressors, and obtaining their piston stroke and reciprocating frequency; S3, obtaining the pressure ratio of each compressor stage, and obtaining the exhaust pressure of the primary and secondary compressors; S4, obtaining the cylinder bore of each compressor stage. This invention can systematically determine the core parameters of each compressor stage, ensuring that the system achieves high exhaust pressure while realizing high displacement output, optimizing internal load distribution, and improving the durability of key moving parts.

Description

Technical Field

[0001] This invention relates to the field of compressor technology, specifically to the design method and compressor system of ultra-high pressure large displacement liquid-driven compressor system. Background Technology

[0002] Hydraulic compressors, with their advantages of compact structure, high pressure, and stable operation, play a crucial role in engineering fields such as military, aerospace, nuclear power, and energy and chemical industries. With the development of related technologies, the market has placed higher demands on the operating pressure (discharge pressure) and discharge flow rate (displacement) of hydraulic compressors, and technological development is moving towards compatibility between ultra-high pressure (e.g., discharge pressure ≥320MPa) and large displacement.

[0003] However, achieving compatibility between ultra-high pressure and large displacement faces significant technical challenges. First, under ultra-high pressure conditions, the compressor's core moving components (such as piston rings and stuffing boxes) endure extremely high mechanical stress and thermal loads, leading to a sharp increase in wear rate and a significant reduction in service life, impacting the long-term operational stability and reliability of the equipment. Second, traditional compressor design methods struggle to effectively coordinate compression ratios at each stage, piston motion parameters, and structural dimensions while ensuring reliability under such extreme conditions, in order to achieve optimal system performance and compactness. Simply increasing the pressure or displacement of a single stage often results in a decrease in the reliability of other components or a reduction in overall efficiency.

[0004] The applicant also developed and applied for a Chinese patent, CN119122779A, entitled "A High-Pressure Liquid-Driven Hydrogen Compressor." By improving the piston rod connection and fit, friction and wear were effectively reduced under high pressure, extending the lifespan of related vulnerable parts. However, this solution primarily focused on optimizing local structures. For the entire compressor system, especially when pursuing the dual goals of "ultra-high pressure" and "large displacement," how to rationally select compressor types at each level, coordinate the design of key operating parameters, and match structural configurations from a system-level design perspective remains a pressing technical challenge. Therefore, the applicant is considering further research and development to achieve stable long-term operation under ultra-high pressure compressor discharge pressure requirements, and to overcome the design challenges of balancing ultra-high pressure discharge with large displacement. Summary of the Invention

[0005] To avoid and overcome the technical problems existing in the prior art, the present invention provides a design method and compressor system for an ultra-high pressure large displacement liquid-driven compressor system, which can systematically determine the core parameters of each stage of the compressor, ensure that the system achieves large displacement output while reaching extremely high exhaust pressure, optimize internal load distribution, and improve the durability of key moving parts.

[0006] To achieve the above objectives, the present invention provides the following technical solution:

[0007] The design method for an ultra-high pressure, large displacement liquid-driven compressor system includes a first-stage compressor, a second-stage compressor, and a third-stage compressor connected in series along the exhaust direction. The parameter design of each compressor includes the following steps:

[0008] S1. Select a primary compressor with a contact-type seal and determine its piston stroke. and frequency of repetition And determine the intake pressure of the compressor system. Maximum exhaust pressure and displacement ;

[0009] S2. Then select the two-stage compressor and the three-stage compressor as intermittent seal type, and the piston stroke of both of them... and according to Obtain, among which, The reduction factor is taken as 0.6~0.7; the piston reciprocating frequency of the second-stage compressor. The reciprocating frequency of the piston in a three-stage compressor ,in, The increment factor is 2 to 4;

[0010] S3. Define the pressure ratio of the first-stage compressor as: The pressure ratio of the secondary compressor is... The pressure ratio of a three-stage compressor ,in, The scaling factor is set to 0.2-0.4 to obtain the pressure ratio of each stage of the compressor, as well as the discharge pressure of the first-stage compressor and the discharge pressure of the second-stage compressor.

[0011] S4. Input the acquired parameters into the cylinder bore model to obtain the cylinder bore of each stage of the compressor.

[0012] As a further aspect of the present invention: the cylinder diameter model in step S4 is specifically as follows:

[0013] ;

[0014] in, The displacement of the compressor system is expressed in Nm³. 3 / h;

[0015] The density of the compressible medium under standard conditions is expressed in kg / m³. 3 ;

[0016] when hour, The diameter of the primary compressor is in mm. The number of effective cylinders in a single-stage compressor, dimensionless; This refers to the piston stroke of a single-stage compressor, expressed in mm. The piston reciprocating frequency of the first-stage compressor is expressed in cycles per minute. The density of the compressed medium under the inlet pressure condition of a single-stage compressor, expressed in kg / m³. 3 ;

[0017] when hour, The diameter of the two-stage compressor is in mm. The effective number of cylinders in the secondary compressor, dimensionless; This refers to the piston stroke of a two-stage compressor, expressed in mm. The piston reciprocating frequency of the two-stage compressor is expressed in cycles per minute. The density of the compressed medium under the inlet pressure condition of the two-stage compressor, in kg / m³. 3 ;

[0018] when hour, The diameter of the three-stage compressor is in mm. The effective number of cylinders in a three-stage compressor, dimensionless; This refers to the piston stroke of a three-stage compressor, expressed in mm. The reciprocating frequency of the piston in a three-stage compressor is expressed in cycles per minute. This refers to the density of the compressed medium under the inlet pressure conditions of a three-stage compressor, expressed in kg / m³. 3 .

[0019] As a further aspect of the present invention: in step S3, the process of obtaining the pressure ratio of each stage of the compressor is specifically as follows:

[0020] First follow Obtain the pressure ratio of the first-stage compressor Then, according to the pressure ratio relationship of each stage of the compressor set in step S3, the pressure ratio of the second stage compressor is obtained sequentially. Pressure ratio of a three-stage compressor .

[0021] As a further aspect of the present invention: in step S3, the process of obtaining the discharge pressure of the primary compressor and the discharge pressure of the secondary compressor is specifically as follows: according to and Obtain them separately, where, This refers to the discharge pressure of the first-stage compressor, expressed in MPa. This is the discharge pressure of the two-stage compressor, expressed in MPa.

[0022] The compressor system is applied to the design method of the ultra-high pressure large displacement liquid-driven compressor system. The first-stage compressor is a double-acting piston compressor with a 1-row 2-cylinder compressor structure. The second-stage and third-stage compressors are both single-acting plunger compressors with a 2-row 4-cylinder compressor structure.

[0023] As a further embodiment of the present invention: the piston compressor includes a power cylinder, and piston cylinders are coaxially mounted on both ends of the power cylinder of the piston compressor, and the power cylinder and piston cylinders are separated by an oil-gas isolation section; the inner cavities of the power cylinder and piston cylinders are respectively sealed and slidably fitted with oil cylinder pistons and air cylinder pistons, and both ends of the piston cylinder are equipped with inlet and outlet valve groups, which are respectively connected to both ends of the inner cavity of the reciprocating stroke of the air cylinder piston; both ends of the power cylinder are equipped with oil cylinder distribution holes, which are respectively connected to both ends of the inner cavity of the reciprocating stroke of the oil cylinder piston; the oil cylinder piston and air cylinder piston are connected by a piston rod, wherein, from both ends to the middle of the outer periphery of the air cylinder piston, a wound spring, a retaining ring, an elastic energy storage sealing ring and an air piston guide ring are coaxially sleeved.

[0024] As a further embodiment of the present invention: the end of the piston rod that extends into the piston cylinder has a stepped shaft structure that is thinner on the outside and thicker on the inside. The cylinder piston is sleeved on the small shaft diameter section of the piston rod, and a locking nut is threaded onto the small shaft diameter section to lock the cylinder piston against the outer end wall of the large shaft diameter section of the piston rod. The outer end of the locking nut has an annular wedge-shaped surface with a gradually decreasing diameter outwards. A locking nut is also threaded onto the small shaft diameter section, and an annular groove is provided on the inner circumference of the locking nut to engage with the annular wedge-shaped surface.

[0025] As a further embodiment of the present invention: the plunger compressor also includes a power cylinder, and the two ends of the power cylinder of the plunger compressor are coaxially mounted with plunger cylinders, and the plunger cylinders and the power cylinder are also separated by an oil-gas isolation section; the piston ends of the power cylinder of the plunger compressor are fixed with plunger rods that slide in the two plunger cylinders in a sealed manner, and the outer ends of the plunger cylinders are all equipped with plunger valve groups, which are connected to the inner cavity at the end of the plunger rod; wherein, along the direction of the plunger rod, the outer periphery of the plunger rod is sequentially fitted with a stuffing gland tightening screw, which together pass through and is fixed on the plunger cylinder with a stuffing ring cover, three sets of stuffing ring bodies, a guide body and a throttling body, and the inner periphery of the throttling body is equipped with a throttling ring, the inner periphery of the guide body is equipped with a stuffing guide ring, and the inner periphery of the stuffing ring body is sequentially equipped with an oblique stuffing ring and a straight stuffing ring along the direction of the plunger rod.

[0026] As a further embodiment of the present invention: an oil-gas isolation cavity is provided on the inner periphery of the oil-gas isolation part, and an oil-gas discharge hole communicating with the bottom of the oil-gas isolation cavity is also provided at the bottom of the oil-gas isolation part. From the inside to the outside, an oil scraper ring, an oil-gas isolation sealing ring and an oil-gas isolation guide ring are provided at both ends of the oil-gas isolation cavity, which are sleeved on the outer periphery of the piston rod or plunger rod.

[0027] As a further embodiment of the present invention: the outer two ends of the cylinder piston are fitted with a Glyd seal ring and a piston guide ring from the outside to the inside, and the two ends of the cylinder piston are locked and fixed to the piston rod or plunger rod by radial set screws.

[0028] Compared with the prior art, the beneficial effects of the present invention are:

[0029] 1. An expanded parametric design method for ultra-high pressure, high-flow compressor systems was developed. The compressor system is divided into a series of sequentially connected primary, secondary, and tertiary compressors with progressively increasing pressure. In the primary compressor, which operates at relatively low pressure, a contact seal (i.e., a piston compressor with piston rings) is selected to achieve better sealing efficiency. In the secondary and tertiary compressors, which face high or even ultra-high pressure operating conditions, a dynamic seal structure with intermittent sealing is chosen to withstand extremely high pressures and ensure the long service life of the liquid-driven compressor.

[0030] Furthermore, by introducing a reduction factor k to control the stroke of the second and third stage compressors, the inertial force and friction path of the high-pressure stage moving parts are reduced while taking into account the intake volume. The reciprocating frequency is gradually increased by an increment factor j to maintain the necessary volumetric flow rate after the stroke is shortened. At the same time, the high-frequency movement also corresponds to the selection of a reciprocating compressor, resulting in lower gas leakage. In addition, the pressure ratio is gradually reduced by a reduction factor M to reasonably distribute the load of each stage and avoid the extreme sealing conditions caused by excessive pressure ratio in the final stage. These factors work together to achieve a system-level balance of high pressure, large displacement, and long service life.

[0031] 2. This application transforms complex design problems into clear steps and models, allowing designers to quickly determine the core dimensions and operating parameters of each stage of the compressor based on the target operating conditions, and to apply them to specific optimized mechanical structures. This greatly improves design efficiency and success rate, and provides a reliable technical path for the development of ultra-high pressure, large displacement liquid-driven compressors.

[0032] 3. The first-stage compressor uses a single-row, two-cylinder, double-acting piston compressor, while the second-stage and third-stage compressors both use a combination structure of two-row, four-cylinder, single-acting plunger compressors. The double-acting piston compressor increases the displacement of the first-stage compressor and helps to reduce the skid-mounted size of the compressor system; while the multi-row plunger compressors operate in parallel in the second and third stages, meeting the displacement requirements under high pressure or even ultra-high pressure, and distributing the load through multiple cylinders, thus improving the stability and reliability of the system operation.

[0033] 4. The cylinder piston of the first-stage compressor adopts a composite sealing and guiding structure. From both ends to the middle, it is sequentially fitted with: a wound snap ring (for positioning), a retaining ring, an elastic energy storage sealing ring (the main seal, which self-tightens under pressure and compensates for wear under elasticity), and a piston guide ring (for support and guidance, reducing lateral wear of the elastic energy storage sealing ring). This combination is particularly suitable for reliable sealing and long-life operation in low-pressure environments of the first-stage compressor.

[0034] 5. The connection between the piston rod and the cylinder piston adopts an anti-loosening structure. On the one hand, the cylinder piston is doubly locked by the locking rear nut and the locking front nut. At the same time, the annular groove on the inner side of the locking front nut and the annular wedge face on the outer end of the locking rear nut form an inclined surface extrusion, generating a strong anti-loosening locking force. This effectively prevents eccentricity under reciprocating impact, reduces lateral wear of the elastic energy storage sealing ring, and further ensures the reliable sealing and long-life operation of the compressor.

[0035] 6. The plunger rod's sealing and guiding system is also a multi-stage combination. From the plunger rod's direction of travel, the components are as follows: a throttling gland (containing a throttling ring for initial pressure reduction and sealing), a guide gland (containing a packing guide ring for support and guidance), three sets of packing ring glands (within each set, beveled packing rings are installed first along the direction of travel, followed by straight packing rings, forming multiple self-tightening seals), and finally, secured by a packing ring gland and packing gland tightening screws. This structure effectively addresses media leakage under ultra-high pressure. Furthermore, the through-locking method of the packing gland tightening screws allows for adjustment or replacement of the packing rings after wear, while also ensuring the coaxiality of the assembly.

[0036] 7. The oil-gas isolation section has an internal oil-gas isolation chamber. The isolation and sealing of the oil-gas isolation section is as follows: at both ends of the chamber, along the axial direction from the inside to the outside, are installed the following in sequence: an oil scraper ring (to scrape oil off the piston rod / plunger rod), an oil-gas isolation sealing ring (the main seal, preventing oil and gas from mixing), and an oil-gas isolation guide ring (to support the rod and reduce wear on the sealing ring). Any leaked trace amounts of oil or gas can be collected in the oil-gas isolation chamber and discharged through the oil-gas discharge port, ensuring the purity of the medium. Attached Figure Description

[0037] Figure 1 This is a schematic diagram of the compressor system structure of the present invention.

[0038] Figure 2 This is a schematic diagram of the piston cylinder in this invention.

[0039] Figure 3 for Figure 2 A magnified schematic diagram of the structure at point A in the middle.

[0040] Figure 4 This is a schematic diagram of the piston cylinder in this invention.

[0041] Figure 5 This is a schematic diagram of the oil-gas isolation section connected to the piston cylinder in this invention.

[0042] Figure 6 This is a schematic diagram of the oil-gas isolation section connected to the plunger cylinder in this invention.

[0043] Figure 7 This is a schematic diagram of the structure of the piston in the power cylinder of the present invention.

[0044] In the diagram: 10, primary compressor; 20, secondary compressor; 30, tertiary compressor;

[0045] 40. Piston cylinder; 41. Piston rod; 411. Front locking nut; 412. Rear locking nut; 42. Cylinder piston; 421. Spiral wound retaining ring; 422. Retaining ring; 423. Elastic energy storage sealing ring; 424. Piston guide ring;

[0046] 50. Plunger cylinder; 51. Throttling box body; 511. Throttling ring; 52. Guide box body; 521. Packing guide ring; 53. Packing ring box body; 531. Straight packing ring; 532. Angled packing ring; 54. Packing ring gland; 55. Packing box tightening screw; 56. Plunger rod;

[0047] 60. Oil-gas isolation section; 61. Oil-gas isolation chamber; 62. Oil-gas discharge port; 63. Oil-gas isolation guide ring; 64. Oil-gas isolation sealing ring; 65. Oil scraper ring;

[0048] 70. Hydraulic cylinder; 71. Hydraulic cylinder piston; 711. Glyd seal ring; 712. Hydraulic piston guide ring; 713. Set screw;

[0049] 80. Cooling water jacket. Detailed Implementation

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

[0051] For ease of understanding, the specific structure and operation of the present invention will be further described below with reference to the accompanying drawings:

[0052] The specific structure of this invention is as follows: Figure 1-7 As shown, it is mainly used to design an intake pressure of Maximum exhaust pressure is Displacement is The system is a liquid-driven compressor system. The system consists of a first-stage compressor 10, a second-stage compressor 20, and a third-stage compressor 30 connected in series.

[0053] Example 1:

[0054] The intake pressure of the liquid-driven compressor MPa, maximum exhaust pressure MPa, compressor displacement Nm 3 Taking N2 as an example, the design of a compressor system is described in detail below. Specifically, the design steps are as follows:

[0055] S1. Determine the parameters of the first-stage compressor (parameter 10):

[0056] Sealing type selection: The working pressure of the first-stage compressor 10 is relatively low, so a contact seal is selected, that is, a piston compressor with piston rings is used to obtain better sealing efficiency.

[0057] Determine the piston stroke and frequency of repetition Based on the initial design, set mm, times / min.

[0058] System-provided parameters: MPa MPa Nm 3 / h.

[0059] S2. Determine the basic motion parameters of the secondary compressor 20 and the tertiary compressor 30:

[0060] Sealing type selection: The two-stage compressor 20 and the three-stage compressor 30 face high pressure or even ultra-high pressure. The service life of the sealing ring of the contact seal is difficult to meet the requirements. Therefore, the dynamic sealing structure of the gap seal is selected, that is, the plunger compressor with plunger and stuffing box seal, so as to withstand extremely high pressure and ensure the long service life requirements of the liquid drive compressor.

[0061] Piston stroke: Considering the higher intake pressure of the second and third stages, sufficient gas can still be drawn in during the short stroke. Therefore, the piston stroke of the second-stage compressor 20 and the third-stage compressor 30 is... and according to Obtain, among which, The reduction factor is 0.6 to 0.7.

[0062] In this embodiment, a reduction factor is used. ,but mm.

[0063] Frequency determination: For dynamic sealing components, higher frequencies result in lower leakage while also maintaining hydraulic system performance. The piston reciprocating frequency of the second-stage compressor 20 is determined according to... Obtain the piston reciprocating frequency of the three-stage compressor 30 according to Obtain, among which, This is an incrementing factor, ranging from 2 to 4.

[0064] In this embodiment, an increasing coefficient is used. ,but times / min, times / min.

[0065] S3. Determine the pressure ratio of each stage and the intermediate pressure:

[0066] To minimize the pressure differential experienced by the high-pressure stage dynamic sealing assembly, the pressure ratio should be reduced as the discharge pressure increases. Therefore, the pressure ratio of the second-stage compressor 20 is set to... The pressure ratio of the three-stage compressor is 30. ,in, The scaling factor is 0.2-0.4.

[0067] In this embodiment, the scaling factor is taken. .

[0068] The total pressure ratio is:

[0069] .

[0070] Based on the pressure ratio relationship of each compressor, the pressure ratio of the first-stage compressor 10 is calculated. The pressure ratio of the second-stage compressor 20 The pressure ratio of the third-stage compressor 30 .

[0071] Calculate the intermediate pressure:

[0072] Discharge pressure of primary compressor 10:

[0073] MPa.

[0074] Discharge pressure of the second-stage compressor 20:

[0075] MPa.

[0076] S4. Calculate the cylinder diameter of each stage of the compressor:

[0077] According to the law of conservation of mass, the total mass of gas medium drawn in by each cylinder per unit time can be calculated. Ignoring gas leakage, the following equation can be derived:

[0078] ;

[0079] Based on the formula conversion, a unified cylinder diameter model formula is obtained:

[0080] ;

[0081] in, The displacement of the compressor system is expressed in Nm³. 3 / h;

[0082] The density of the compressible medium under standard conditions is expressed in kg / m³. 3 ;

[0083] when hour, The diameter of the primary compressor 10 is in mm. The effective number of cylinders in the primary compressor 10 is dimensionless. The piston stroke of the first-stage compressor 10 is in mm. The reciprocating frequency of the piston in the first-stage compressor 10 is expressed in cycles per minute. The density of the compressed medium under the inlet pressure condition of a single-stage compressor is expressed in kg / m³. 3 ;

[0084] when hour, The diameter of the secondary compressor 20 is in mm; The number of effective cylinders in the 20-stage compressor is dimensionless. The piston stroke of the two-stage compressor 20 is in mm. The piston reciprocating frequency of the secondary compressor 20 is expressed in cycles per minute. This refers to the density of the compressed medium under an intake pressure of 20 kg / m³ for a two-stage compressor. 3 ;

[0085] when hour, The diameter of the 30-stage compressor is in mm. The effective number of cylinders in a three-stage compressor 30 is dimensionless. The piston stroke of a three-stage compressor (30) is in mm. The reciprocating frequency of the piston in a three-stage compressor 30 is expressed in cycles per minute. The density of the compressed medium under an intake pressure of 30 ppm for a three-stage compressor, expressed in kg / m³. 3 It can be calculated using the gas law.

[0086] In the parameters, kg / m 3 , Nm 3 / h.

[0087] First-stage compressor 10: Inlet pressure MPa, assuming the compression medium is N2, at this intake pressure kg / m 3 The first-stage compressor 10 uses a double-acting reciprocating compressor, with an effective number of cycles. =4 (double-acting, working at both ends of each cylinder stroke).

[0088] but: 75mm.

[0089] Secondary compressor 20: Intake pressure is the same as the discharge pressure of primary compressor 10. MPa, at this intake pressure kg / m 3 It adopts a single-acting, 2-row, 4-cylinder configuration. =4.

[0090] but 67mm.

[0091] Three-stage compressor 30: Intake pressure is the same as the discharge pressure of two-stage compressor 20. MPa, at this intake pressure kg / m 3 It adopts a single-acting, 2-row, 4-cylinder configuration. =4.

[0092] but 53mm.

[0093] Through the above calculations, the core structural parameters (cylinder diameter, stroke, frequency) and operating parameters (pressure ratio) of each level of compressor can be obtained to meet the requirements of 320MPa ultra-high pressure exhaust and 500Nm³ / h large displacement.

[0094] Example 2: Compressor System Structure

[0095] Figures 1-7 The specific compressor system structure constructed using the design method of Example 1 is illustrated. The system's first-stage compressor 10 is a piston type, while the second-stage compressor 20 and the third-stage compressor 30 are plunger types, and their specific construction is as follows:

[0096] like Figure 1 As shown, the single-stage reciprocating compressor includes a power cylinder 70, whose left and right ends are coaxially connected to piston cylinders 40 via oil-gas isolation parts 60 (in actual implementation, a cooling water jacket 80 is provided on the outside of the piston cylinder 40 for rapid cooling of the cylinder body). The cylinder piston 71 inside the power cylinder 70 is rigidly connected to the cylinder piston 42 inside the piston cylinder 40 via a piston rod 41. Oil distribution holes are provided at both ends of the power cylinder 70. Figure 2 and Figure 4 As shown), the hydraulic cylinder piston 71 reciprocates via hydraulic oil, which in turn drives the pneumatic cylinder piston 42. The piston cylinder 40 has a one-way intake valve and a one-way exhaust valve at both ends (collectively referred to as the intake and exhaust valve assembly, such as...). Figure 2 As shown in the figure, dual-action compression is achieved.

[0097] like Figure 2 and Figure 3 As shown, the cylinder piston 42 employs a composite sealing and guiding structure. From both ends towards the center, the following components are sequentially fitted: a wound snap ring 421 (for positioning), a retaining ring 422, an elastic energy storage sealing ring 423 (the main seal, which self-tightens under pressure and compensates for wear under elasticity), and a piston guide ring 424 (for support and guidance, reducing lateral wear of the elastic energy storage sealing ring 423). This combination is particularly suitable for reliable sealing and long-life operation in low-pressure environments of the first-stage compressor 10.

[0098] like Figure 2 and Figure 3 As shown, the connection between the piston rod 41 and the cylinder piston 42 employs an anti-loosening structure. The end of the piston rod 41 is a stepped shaft with a thinner outer diameter and a thicker inner diameter. The cylinder piston 42 is fitted onto the smaller diameter section and is initially locked by the locking nut 412, which abuts against the outer end face of the larger diameter section. The locking nut 411 is also threaded onto the smaller diameter section. By screwing it in, the annular groove on its inner side and the annular wedge-shaped surface on the outer end of the locking nut 412 form a sloping extrusion, generating a strong anti-loosening locking force. This effectively prevents the cylinder piston 42 from becoming eccentric under reciprocating impact, reduces lateral wear of the elastic energy storage sealing ring 423, and further ensures the reliable sealing and long-life operation of the first-stage compressor 10.

[0099] like Figure 4As shown, the two-stage compressor 20 and the three-stage compressor 30 both employ plunger compressors, which also include a power cylinder 70. Both ends are connected to the plunger cylinders 50 via oil-gas isolation sections 60 (in actual implementation, the plunger cylinders 50 are also equipped with cooling water jackets 80 for rapid cooling of the cylinder body). The difference lies in that the plunger rods 56, fixed at both ends of the piston 71 of the power cylinder 70, enter the plunger cylinders 50 on both sides, forming a sliding reciprocating clearance seal with the plunger rods 56 and the plunger cylinders 50. A plunger valve assembly (including one-way intake and exhaust valves, such as...) is installed at the outer end of the plunger cylinders 50. Figure 4 As shown in the figure, single-action compression is achieved.

[0100] like Figure 4 As shown, the sealing and guiding system of the plunger rod 56 is also a multi-stage combination. From the direction of the plunger rod 56's movement (i.e., from the sliding direction from the power cylinder 70 to the plunger cylinder 50), the components are as follows: a throttling gland 51 (containing a throttling ring 511 for initial pressure reduction and sealing), a guide gland 52 (containing a packing guide ring 521 for support and guidance), three sets of packing ring glands 53 (within each set, a beveled packing ring 532 is installed first along the movement direction, followed by a straight packing ring 531, forming multiple self-tightening seals), and finally, it is tightened and fixed by the packing ring cap 54 and the packing gland locking screws 55. This structure can effectively cope with medium leakage under ultra-high pressure. Furthermore, the through-locking method of the packing gland locking screws 55 allows for adjustment or replacement of the packing rings after wear, while also ensuring the coaxiality of the assembly.

[0101] like Figure 5 and Figure 6 As shown, the oil-gas isolation section 60 has an internal oil-gas isolation chamber 61 for isolating gas and hydraulic oil between the power cylinder 70 and the plunger cylinder 50 or piston cylinder 40, and temporarily storing gas and hydraulic oil in case of leakage. It has an oil-gas discharge hole 62 at its bottom. Specifically, the isolation seal of the oil-gas isolation section 60 is as follows: at both ends of the chamber, along the axial direction from the inside to the outside, are installed the following: an oil scraper ring 65 (to scrape oil off the piston rod 41 / plunger rod 56), an oil-gas isolation sealing ring 64 (the main seal, preventing oil and gas from mixing), and an oil-gas isolation guide ring 63 (a support member, reducing wear on the sealing ring). Any leaked trace amounts of oil or gas can be collected in the oil-gas isolation chamber 61 and discharged through the oil-gas discharge hole 62, ensuring the purity of the medium.

[0102] like Figure 7 As shown, the cylinder piston 71 is sealed with a Glyd ring 711, which has good sealing performance and high pressure resistance. It is supported by an oil piston guide ring 712 on the outer side. The cylinder piston 71 is fixed to the piston rod 41 or plunger rod 56 by radial set screws 713, ensuring a reliable connection.

[0103] In summary, the design methodology of this application complements the specific compressor system structure. The design methodology provides the theoretical parameter basis and configuration strategy for achieving the ultra-high pressure and large displacement target, while the innovative designs of the specific piston / plunger compressor structure, sealing system, and connection method provide solid physical support for the reliable and stable realization of these theoretical parameters under extreme operating conditions.

[0104] Of course, those skilled in the art will recognize that the present invention is not limited to the details of the exemplary embodiments described above, but also includes the same or similar structures that can be implemented in other specific forms without departing from the spirit or essential characteristics of the invention. Therefore, the embodiments should be considered illustrative and non-limiting in all respects, and the scope of the invention is defined by the appended claims rather than the foregoing description. Thus, all variations falling within the meaning and scope of equivalents of the claims are intended to be included within the present invention. No reference numerals in the claims should be construed as limiting the scope of the claims.

[0105] Furthermore, it should be understood that although this specification describes embodiments, not every embodiment contains only one independent technical solution. This narrative style is merely for clarity. Those skilled in the art should consider the specification as a whole, and the technical solutions in each embodiment can also be appropriately combined to form other embodiments that can be understood by those skilled in the art.

[0106] The technologies, shapes, and structures not described in detail in this invention are all known technologies.

Claims

1. A design method for an ultra-high pressure, large displacement liquid-driven compressor system, characterized in that, The system includes a primary compressor (10), a secondary compressor (20), and a tertiary compressor (30) connected in series along the exhaust direction. The parameter design of each compressor includes the following steps: S1. Select a primary compressor (10) with a contact-type seal and determine its piston stroke. and frequency of repetition And determine the intake pressure of the compressor system. Maximum exhaust pressure and displacement ; S2. Then select the secondary compressor (20) and the tertiary compressor (30) as intermittent seal type, and the piston stroke of both is... and according to Obtain, among which, The reduction factor is taken as 0.6~0.7; the piston reciprocating frequency of the secondary compressor (20) The reciprocating frequency of the piston in the three-stage compressor (30) ,in, The increment factor is 2 to 4; S3. Define the pressure ratio of the first-stage compressor (10) as follows: The pressure ratio of the secondary compressor (20) is... The pressure ratio of the three-stage compressor (30) ,in, The scaling factor is set to 0.2-0.

4. The pressure ratio of each stage of the compressor is obtained, and the discharge pressure of the first stage compressor (10) and the discharge pressure of the second stage compressor (20) are obtained. S4. Input the acquired parameters into the cylinder bore model to obtain the cylinder bore of each stage of the compressor.

2. The design method for the ultra-high pressure, large displacement liquid-driven compressor system according to claim 1, characterized in that, The cylinder bore model in step S4 is specifically as follows: ； in, The displacement of the compressor system, in Nm³. 3 / h; The density of the compressible medium under standard conditions is expressed in kg / m³. 3 ; when hour, The diameter of the primary compressor (10) is in mm; The effective number of cylinders in the primary compressor (10) is dimensionless; The piston stroke of the first-stage compressor (10) is in mm; The piston reciprocating frequency of the first-stage compressor (10) is expressed in cycles / min. The density of the compressed medium under the intake pressure condition of the first-stage compressor (10), in kg / m³. 3 ; when hour, The diameter of the secondary compressor (20) is in mm; The effective number of cylinders in the secondary compressor (20) is dimensionless; The piston stroke of the two-stage compressor (20) is in mm; The reciprocating frequency of the piston of the secondary compressor (20) is expressed in cycles / min. The density of the compressed medium under the intake pressure condition of the two-stage compressor (20), in kg / m³. 3 ; when hour, The diameter of the three-stage compressor (30) is in mm; The effective number of cylinders in a three-stage compressor (30) is dimensionless; The piston stroke of the three-stage compressor (30) is in mm; The reciprocating frequency of the piston in the three-stage compressor (30) is expressed in cycles / min. The density of the compressed medium under the inlet pressure condition of the three-stage compressor (30), in kg / m³. 3 .

3. The design method for an ultra-high pressure, large displacement hydraulically driven compressor system according to claim 1, characterized in that, In step S3, the process of obtaining the pressure ratio of each stage of the compressor is as follows: First follow Obtain the pressure ratio of the first-stage compressor (10) Then, according to the pressure ratio relationship of each stage of the compressor set in step S3, the pressure ratio of the second stage compressor (20) is obtained sequentially. Pressure ratio of the three-stage compressor (30) .

4. The design method for an ultra-high pressure, large displacement hydraulically driven compressor system according to claim 1, characterized in that, In step S3, the process of obtaining the discharge pressure of the primary compressor (10) and the discharge pressure of the secondary compressor (20) is specifically as follows: according to and Obtain them separately, where, The discharge pressure of the first-stage compressor (10) is expressed in MPa. The discharge pressure of the secondary compressor (20) is expressed in MPa.

5. A compressor system, wherein the compressor system is applied to the design method of the ultra-high pressure large displacement liquid-driven compressor system as described in any one of claims 1-4, characterized in that, The first-stage compressor (10) is a double-acting piston compressor with a 1-row 2-cylinder compressor structure; the second-stage compressor (20) and the third-stage compressor (30) are both single-acting plunger compressors with a 2-row 4-cylinder compressor structure.

6. The compressor system according to claim 5, characterized in that, The piston compressor includes a power cylinder (70), and piston cylinders (40) are coaxially mounted on both ends of the power cylinder (70), and the power cylinder (70) and piston cylinders (40) are separated by an oil-gas isolation section (60); the inner cavities of the power cylinder (70) and piston cylinders (40) are respectively sealed and slidably fitted with a cylinder piston (71) and a cylinder piston (42), and both ends of the piston cylinder (40) are equipped with inlet and outlet valve groups, which are respectively connected to the cylinder piston (42). 2) The two ends of the reciprocating stroke inner cavity are connected. Both ends of the power cylinder (70) are equipped with cylinder oil distribution holes. The two sets of cylinder oil distribution holes are connected to the two ends of the reciprocating stroke inner cavity of the cylinder piston (71). The cylinder piston (71) and the air cylinder piston (42) are connected by the piston rod (41). The outer periphery of the air cylinder piston (42) is coaxially fitted with a spiral snap ring (421), a retaining ring (422), an elastic energy storage sealing ring (423), and an air piston guide ring (424) from both ends to the middle.

7. The compressor system according to claim 6, characterized in that, The end of the piston rod (41) that extends into the piston cylinder (40) has a stepped shaft structure that is thinner on the outside and thicker on the inside. The cylinder piston (42) is sleeved on the small shaft diameter section of the piston rod (41), and a locking nut (412) is threaded onto the small shaft diameter section to lock the cylinder piston (42) against the outer end wall of the large shaft diameter section of the piston rod (41). The outer end of the locking nut (412) has an annular wedge-shaped surface with a gradually decreasing diameter. A locking nut (411) is also threaded onto the small shaft diameter section. The inner circumference of the locking nut (411) is provided with an annular groove that wedges with the annular wedge-shaped surface.

8. The compressor system according to claim 6, characterized in that, The plunger compressor also includes a power cylinder (70). A plunger cylinder (50) is coaxially mounted at both ends of the power cylinder (70) of the plunger compressor, and the plunger cylinder (50) and the power cylinder (70) are separated by an oil-gas isolation section (60). The piston (71) of the power cylinder (70) of the plunger compressor has plunger rods (56) fixed at both ends, which slide and seal within the two plunger cylinders (50). Inlet and outlet valve assemblies are installed on the cylinder heads that connect to the end faces of the plunger cylinders (50), and these valve assemblies communicate with the inner cavity at the end of the plunger rod (56). Along the plunger rod (56)... 6) In the direction of progress, the outer periphery of the plunger rod (56) is sequentially fitted with a stuffing box tightening screw (55) and a stuffing ring cover (54), three sets of stuffing ring bodies (53), guide bodies (52) and throttling bodies (51) fixed on the plunger cylinder (50). The inner periphery of the throttling body (51) is fitted with a throttling ring (511), the inner periphery of the guide body (52) is fitted with a stuffing guide ring (521), and the inner periphery of the stuffing ring body (53) is sequentially fitted with a beveled stuffing ring (532) and a straight stuffing ring (531) along the direction of progress of the plunger rod (56).

9. The compressor system according to claim 8, characterized in that, The inner circumference of the oil-gas isolation section (60) is provided with an oil-gas isolation cavity (61), and the bottom of the oil-gas isolation section (60) is also provided with an oil-gas discharge hole (62) that connects to the bottom of the oil-gas isolation cavity (61). The two ends of the oil-gas isolation cavity (61) are provided with an oil scraper ring (65), an oil-gas isolation sealing ring (64) and an oil-gas isolation guide ring (63) respectively, which are sleeved on the outer circumference of the piston rod (41) or the plunger rod (56).

10. The compressor system according to claim 8, characterized in that, The outer two ends of the cylinder piston (71) are fitted with a Glyd seal ring (711) and a piston guide ring (712) from the outside to the inside. The two ends of the cylinder piston (71) are locked and fixed to the piston rod (41) or plunger rod (56) by radial set screws (713).

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