A diaphragm compressor with a two-stage counter-acting plunger pump structure and a design method thereof

Abstract

A diaphragm compressor with a two-stage reciprocating plunger pump structure and a design method thereof, the diaphragm compressor comprising a first-stage plunger pump and a second-stage plunger pump, the first-stage plunger pump and the second-stage plunger pump being respectively installed on two symmetrical sides of a compressor end cover; hydraulic oil sealing is performed between plunger pump bodies of the first-stage plunger pump and the second-stage plunger pump and the compressor end cover, between the plunger pump bodies and plunger pump bushings, between the plunger pump bodies and plunger pump covers, and between the plunger pump bushings and internal plunger pump plungers; the first-stage plunger pump and the second-stage plunger pump are driven by eccentric wheel structures and are jointly acted on by reaction forces of plunger pump springs and frictional forces of the plunger pump plungers to realize processes of oil feeding, pressure increasing, oil discharging, and pressure decreasing. The present application can determine diameters and strokes of the plunger pump plungers, determine minimum spring forces required by the plunger pump springs, and determine optimal installation angle ranges of the two-stage plunger pumps to provide design schemes, performance prediction, and optimization directions for improving service life and reliability of the plunger pump.

Description

TECHNICAL FIELD

[0001] The application belongs to the technical field of diaphragm compressors, and particularly relates to a two-stage counter-driven plunger pump structure diaphragm compressor and a design method thereof. BACKGROUND

[0002] A diaphragm compressor is a positive displacement compressor that separates the oil side and the gas side by a diaphragm. Its working principle is that the high-pressure oil liquid is pressurized and released by driving the crankshaft connecting rod to drive the piston to reciprocate, and then the gas is compressed and discharged by the high-pressure oil liquid pushing the diaphragm. The diaphragm compressor is widely used in the hydrogen energy industry due to its good sealing performance and large pressure ratio, and is particularly important in hydrogenation stations. However, the diaphragm life, volumetric efficiency and diaphragm head strength are key and difficult problems in the development process of the diaphragm compressor.

[0003] The oil supplementing, hydraulic and lubricating systems of the diaphragm compressor directly affect the operation reliability of the whole machine. The hydraulic oil in the diaphragm head assembly is driven by the crank connecting rod mechanism to drive the piston, which directly affects the gas side pressure and gas displacement, and directly acts on the diaphragm to cause its deflection and deformation, and directly acts on the diaphragm head assembly to affect its strength reliability. When the diaphragm compressor is running, high-pressure hydraulic oil will inevitably leak through the piston ring, which will directly lead to an increase in the clearance of the diaphragm compressor and a decrease in the gas volume, and even loss of the gas compression capacity. Therefore, the diaphragm compressor needs to be provided with an oil supplementing plunger pump to supplement the hydraulic oil in the diaphragm head. The oil supplementing plunger pump is generally driven by the eccentric wheel of the crankshaft shaft head, and forms a certain phase difference with the piston in the crankcase to supplement the oil. The traditional diaphragm compressor generally sets the phase difference to 0° or 180°, that is, the plunger pump performs plunger pump oil supplementing operation during the suction or discharge stroke of the diaphragm compressor. The patent with publication number CN109469603B proposes a high-pressure diaphragm compressor and an oil supplementing system and method thereof. An overflow valve passage is connected between the oil supplementing plunger pump and the oil supplementing one-way valve to realize oil supplementing pressure control during the suction or discharge stroke of the diaphragm compressor, thereby realizing the regulation and control of the plunger pump oil supplementing amount.

[0004] In the oil supplementing system, the plunger pump stroke volume is generally designed to be greater than the hydraulic oil leakage in the oil cylinder. However, excessive plunger pump oil supplementing will affect the suction volume and cause insufficient single rotation suction volume. The oil supplementing amount is one of the key problems to ensure the stable operation of the diaphragm compressor. How to reasonably set the plunger pump stroke volume is a difficult problem in the current diaphragm compressor design stage. SUMMARY

[0005] The application aims to solve the problems in the prior art, and provides a two-stage counter-driven plunger pump structure diaphragm compressor and a design method thereof, which improves the service life of the plunger pump of the diaphragm compressor and the reliability of the whole machine of the diaphragm compressor.

[0006] To achieve the above object, the present application has the following technical solutions:

[0007] A diaphragm compressor with a two-stage reciprocating plunger pump structure, comprising a compressor end cover, a first-stage plunger pump, a second-stage plunger pump, plunger pump connecting bolts, a first-stage eccentric structure, and a second-stage eccentric structure; the first-stage plunger pump and the second-stage plunger pump are respectively installed on the symmetrical two sides of the compressor end cover through the plunger pump connecting bolts to achieve power balance; the plunger pump body of the first-stage plunger pump and the second-stage plunger pump is sealed from the compressor end cover, the plunger pump body and the plunger pump bushing, the plunger pump body and the plunger pump cover, and the plunger pump bushing and the internal plunger pump plunger through sealing rings for hydraulic oil sealing; the first-stage plunger pump and the second-stage plunger pump are driven by the first-stage eccentric structure and the second-stage eccentric structure and are jointly acted on by the reaction force of the plunger pump spring and the friction force of the plunger pump plunger to realize the processes of oil intake, pressure increase, oil discharge, and pressure decrease of the plunger pump.

[0008] As a preferred solution, the plunger pump body of the first-stage plunger pump and the second-stage plunger pump is sealed from the compressor end cover through a sealing ring in the fourth sealing groove of the plunger pump body, the plunger pump body and the plunger pump bushing are sealed through sealing rings in the second sealing groove and the third sealing groove of the plunger pump body, the plunger pump cover and the plunger pump body are sealed through a sealing ring in the first sealing groove of the plunger pump body, and the plunger pump bushing and the plunger pump plunger are sealed through sealing rings in multiple plunger sealing grooves.

[0009] As a preferred solution, the first-stage plunger pump and the second-stage plunger pump are driven by the first-stage eccentric structure and the second-stage eccentric structure, the plunger pump plunger moves in the compression direction of the plunger pump spring, the volume of the plunger pump oil cavity decreases in this process, the process of pressure increase and oil discharge of the hydraulic oil in the plunger pump oil cavity is realized, and the plunger pump spring provides a reaction force against the driving force of the eccentric and the friction force of the plunger pump plunger. When the plunger pump plunger moves to the top dead center position, the plunger pump plunger is driven by the plunger pump spring to move to the bottom dead center position against the reaction force of the eccentric and the friction force of the plunger pump plunger. In this process, the volume of the plunger pump oil cavity increases, the process of pressure decrease and oil intake of the hydraulic oil in the plunger pump oil cavity is realized, and the plunger pump plunger reaches the bottom dead center position.

[0010] As a preferred solution, the volume of the volume cavity of the first-stage plunger pump and the second-stage plunger pump is greater than the leakage amount of the diaphragm compressor piston leakage passage gap L xl ;

[0011] The leakage amount of the diaphragm compressor piston leakage passage gap L xl is calculated according to the following formula:

[0012]

[0013] wherein Q l is the leakage flow of the diaphragm compressor piston leakage passage gap L xl ; d hs is the diameter of the diaphragm compressor piston; δ is the gap of the diaphragm compressor piston leakage passage gap L xl ; μ is the viscosity of the hydraulic oil; l is the length of the diaphragm compressor piston leakage passage gap L xl ; Δp is the pressure difference on both sides of the diaphragm compressor piston leakage passage gap L xl ; p oil is the maximum value of the diaphragm compressor oil pressure; p air is the atmospheric pressure; and m is the number of sealing rings on the diaphragm compressor piston.

[0014] As a preferred solution, the volume cavity volume of the primary piston pump and the secondary piston pump is calculated according to the following formula:

[0015]

[0016] wherein m 柱塞 is the oil supplement of the primary piston pump or the secondary piston pump; m l is the leakage mass of the diaphragm compressor piston leakage passage gap; χ m is the design coefficient, greater than 1; ρ oil is the density of the hydraulic oil; d zs is the diameter of the piston pump piston; L zs is the stroke of the piston pump piston.

[0017] As a preferred solution, the reaction force of the piston pump spring is calculated according to the following formula:

[0018]

[0019] wherein F spring,min is the minimum spring force of the piston pump spring at the bottom dead center position; E is the elastic modulus of the sealing ring in the piston sealing groove; H is the wire diameter of the sealing ring in the piston sealing groove; h is the groove depth of the sealing ring in the piston sealing groove; D is the inner diameter of the sealing ring in the piston sealing groove; b is the sealing groove width of the piston sealing groove; μ is the Poisson's ratio of the sealing ring in the piston sealing groove; l is the characteristic length of the piston pump piston in the axial direction contacting the leakage hydraulic oil; ν is the kinematic viscosity of the hydraulic oil; ρ is the density of the hydraulic oil; u is the moving speed of the piston pump piston; and s is the circumferential gap of the piston pump piston.

[0020] As a preferred scheme, the reciprocating movement direction of the diaphragm compressor piston and the reciprocating movement direction of the plunger pump plunger exist a phase difference, the phase difference is determined by the relative high pressure and the relative low pressure of the diaphragm compressor; the relative high pressure and the relative low pressure are calculated as follows:

[0021]

[0022] In the formula, p relative,high is the relative high pressure; χ h is the relative high pressure coefficient; p oil,max is the maximum oil pressure; χ l is the relative low pressure coefficient; p stage,1 is the first stage instantaneous oil pressure of the diaphragm compressor; p stage,2 is the second stage instantaneous oil pressure of the diaphragm compressor; t s is the diaphragm compressor overflow valve oil supplement start phase angle, t e is the diaphragm compressor overflow valve oil supplement end phase angle.

[0023] As a preferred scheme, the first stage plunger pump and the second stage plunger pump are symmetrically installed at 180°, by selecting the relative high pressure coefficient χ h and the relative low pressure coefficient χ l , the diaphragm compressor overflow valve oil supplement start phase angle t s and the diaphragm compressor overflow valve oil supplement end phase angle t e satisfy:

[0024]

[0025] As a preferred scheme, the relative movement direction of the first stage plunger pump and the second stage plunger pump and the relative movement direction of the double stage counteracting diaphragm compressor piston movement exist a phase angle , which meets the following relationship:

[0026]

[0027] The first stage of the diaphragm compressor inhales air within , and compresses exhaust within , the best time for the plunger pump to start oil supplement to the membrane head is at [t1, t s ], and the best time for the plunger pump to start oil supplement is at [t e , T]∪[0, t1].

[0028] A design method of a double stage counteracting plunger pump structure diaphragm compressor as described, comprising the following steps:

[0029] Define the in-cylinder oil pressure p stage , then the dimensionless oil pressure is described as Here, the maximum oil pressure poil,max The nominal discharge pressure is designed for the diaphragm compressor; the nominal dimensionless oil pressure is described as follows: The dimensionless oil pressure of the diaphragm compressor as a function of crank angle was obtained through thermodynamic process simulation calculations. The variation curve of the relative high pressure coefficient χ is selected by... h With relative low pressure coefficient χ l This causes the diaphragm compressor overflow valve to begin replenishing oil at a phase angle t. s Phase angle t of the diaphragm compressor overflow valve oil replenishment end e satisfy:

[0030]

[0031] Determine the phase angle between the relative motion directions of the first-stage and second-stage piston pumps and the relative motion direction of the pistons in the two-stage opposing diaphragm compressor. Determine the installation angle range for the primary and secondary plunger pumps.

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

[0033] This invention relates to a two-stage counteracting plunger pump structure diaphragm compressor. The first and second stage plunger pumps, driven by eccentric wheel structures, respectively, and acting under the combined action of the plunger pump spring's reaction force and the plunger's friction, achieve the processes of oil inlet, pressurization, oil discharge, and depressurization. This allows for the determination of the plunger diameter and stroke, the minimum spring force required for the plunger pump springs to match the overall structural dimensions of the designed plunger pump, and the determination of the optimal installation angle range for the two-stage plunger pumps, thereby minimizing the working pressure and maximizing the oil replenishment. This provides structural design solutions, performance predictions, and optimization directions for improving the service life and reliability of the plunger pump. The two-stage counteracting plunger pump structure diaphragm compressor designed in this invention has a simple structure, small size, flexible and convenient use, and is suitable for serial production or customization.

[0034] Furthermore, the two-stage counteracting plunger pump structure diaphragm compressor of the present invention uses a simplified hydraulic oil leakage gap channel model to quantitatively describe and calculate the oil cavity leakage of the diaphragm compressor, and uses a correction coefficient method to quantitatively describe and calculate the plunger pump cavity stroke volume, replacing the method of determining the plunger pump stroke volume by empirical ratio of oil cavity volume.

[0035] Furthermore, this invention quantitatively describes and calculates the minimum spring force required for the operation of the plunger pump, taking into account the frictional force of the plunger pump sealing ring and the viscous frictional force of the hydraulic oil, thereby optimizing the selection of the plunger pump spring.

[0036] Further, the application quantitatively describes and calculates the optimal installation phase difference of the double-stage diaphragm compressor plunger pump, determines the optimal oil injection time range and the oil supplement starting time of the plunger pump, and is favorable for improving the service life of the diaphragm compressor plunger pump and the reliability of the diaphragm compressor. BRIEF DESCRIPTION OF DRAWINGS

[0037] In order to more clearly illustrate the technical solutions in the specific embodiments or the prior art, the drawings needed in the following description of the specific embodiments or the prior art are briefly introduced. Obviously, the drawings in the following description are some embodiments of the application, and other drawings can be obtained by those skilled in the art without any creative effort.

[0038] Figure 1 It is a schematic diagram of the overall cross-sectional structure of the double-stage plunger pump and the end cover of the double-stage opposed diaphragm compressor of the embodiment of the application.

[0039] Figure 2 It is an enlarged schematic diagram of the cross-sectional structure of the plunger pump of the embodiment of the application.

[0040] Figure 3 It is a schematic diagram of the cross-sectional structure of the piston and cylinder sleeve assembly and the gap leakage passage of the embodiment of the application.

[0041] Figure 4 It is a schematic diagram of the optimal oil supplement time of the plunger pump of the embodiment of the application.

[0042] Figure 5 It is a characteristic change curve of the oil pressure and the gas pressure of the first stage of the diaphragm compressor of the embodiment of the application.

[0043] Figure 6 It is a dimensionless oil pressure change curve with crank angle obtained by AMESim simulation calculation of the double-stage diaphragm compressor of the embodiment of the application.

[0044] In the drawings: A-compressor end cover; B1-first stage plunger pump; B2-second stage plunger pump; C-plunger pump connecting bolt; D1-first stage eccentric structure; D2-second stage eccentric structure; E-piston cylinder sleeve; F-piston; L xl - diaphragm compressor piston leakage passage gap; 1-plunger pump body; 11-plunger pump body first sealing groove; 12-plunger pump body second sealing groove; 13-plunger pump body third sealing groove; 14-plunger pump body fourth sealing groove; 15-plunger pump body oil passage; 2-plunger pump cover; 3-plunger pump bushing; 4-plunger pump spring; 5-plunger pump plunger; 51-plunger sealing groove; 6-plunger pump oil inlet; 7-plunger pump oil outlet. DETAILED DESCRIPTION

[0045] The technical solutions of the present application will be described clearly and completely below in connection with the embodiments. Obviously, the described embodiments are only part of the embodiments of the present application, rather than all the embodiments. Based on the embodiments of the present application, all the other embodiments obtained by those skilled in the art without creative work fall within the scope of the present application.

[0046] Please refer to Figure 1 The embodiment of the present application provides a diaphragm compressor with a two-stage counteracting plunger pump structure, which comprises a compressor end cover A, a first-stage plunger pump B1, a second-stage plunger pump B2, plunger pump connecting bolts C, a first-stage eccentric structure D1 and a second-stage eccentric structure D2. The first-stage plunger pump B1 and the second-stage plunger pump B2 are respectively installed on the symmetrical two sides of the compressor end cover A through the plunger pump connecting bolts C to achieve power balance. The plunger pump body 1 of the first-stage plunger pump B1 and the second-stage plunger pump B2 is sealed from hydraulic oil through a sealing ring between the plunger pump body 1 and the compressor end cover A, between the plunger pump body 1 and the plunger pump bushing 3, between the plunger pump body 1 and the plunger pump cover 2, and between the plunger pump bushing 3 and the plunger pump plunger 5 inside. The first-stage plunger pump B1 and the second-stage plunger pump B2 are driven by the first-stage eccentric structure D1 and the second-stage eccentric structure D2 respectively, and are jointly acted on by the reaction force of the plunger pump spring 4 and the friction force of the plunger pump plunger 5 to realize the processes of oil feeding, pressure increasing, oil discharging and pressure decreasing of the plunger pump.

[0047] Please refer to Figure 2 In a possible implementation, the plunger pump body 1 of the first-stage plunger pump B1 and the second-stage plunger pump B2 is sealed from hydraulic oil through a sealing ring in the fourth sealing groove 14 of the plunger pump body, the plunger pump body 1 and the plunger pump bushing 3 are sealed from hydraulic oil through sealing rings in the second sealing groove 12 and the third sealing groove 13 of the plunger pump body, the plunger pump cover 2 and the plunger pump body 1 are sealed from hydraulic oil through a sealing ring in the first sealing groove 11 of the plunger pump body, and the plunger pump bushing 3 and the plunger pump plunger 5 are sealed from hydraulic oil through sealing rings in the six plunger sealing grooves 51.

[0048] Furthermore, in this embodiment of the invention, the primary plunger pump B1 and the secondary plunger pump B2 are driven by the primary eccentric wheel structure D1 and the secondary eccentric wheel structure D2, respectively. The plunger 5 of the plunger pump moves in the compression direction of the plunger pump spring 4. During this process, the volume of the plunger pump oil chamber decreases, realizing the pressurization and discharge process of the hydraulic oil in the plunger pump oil chamber. At the same time, the plunger pump spring 4 provides a reaction force to counteract the driving force of the eccentric wheel and the friction force of the plunger pump plunger 5. When the plunger pump plunger 5 moves to the top dead center position, under the driving action of the plunger pump spring 4, the plunger pump plunger 5 counteracts the reaction force of the eccentric wheel and the friction force of the plunger pump plunger 5, and pushes the plunger pump plunger 5 to move to the bottom dead center position. During this process, the volume of the plunger pump oil chamber increases, realizing the depressurization and oil intake process of the hydraulic oil in the plunger pump oil chamber, until the plunger pump plunger 5 reaches the bottom dead center position. This process is repeated to achieve the oil inlet, pressurization, oil discharge, and depressurization processes of the primary plunger pump B1 and the secondary plunger pump B2.

[0049] In one possible implementation, the volume of the primary plunger pump B1 and the secondary plunger pump B2 in this embodiment of the invention is greater than the piston leakage passage gap L of the diaphragm compressor. xl If we describe the leakage channel as a concentric cylindrical toroidal gap, then the diaphragm compressor piston leakage channel gap L... xl The leakage amount is calculated according to the following formula:

[0050]

[0051] In the formula, Q l For the piston leakage passage clearance L of the diaphragm compressor xl Leakage flow rate, in kg·s -1 ;d hs δ represents the diameter of the diaphragm compressor piston F, in meters (m); δ represents the leakage passage clearance L of the diaphragm compressor piston. xl The gap, in meters;

[0052] μ represents the hydraulic oil viscosity, measured in Pa·s; l represents the piston leakage passage clearance of the diaphragm compressor. xl The length, in meters (m);

[0053] Δp is the piston leakage passage clearance L of the diaphragm compressor. xl The pressure difference across the two sides, in Pa; p oil This represents the maximum oil pressure of the diaphragm compressor, in Pa; p air Atmospheric pressure, in Pa; m is the number of sealing rings on piston F of the diaphragm compressor, in units of one.

[0054] Furthermore, the volumes of the primary plunger pump B1 and the secondary plunger pump B2 are calculated according to the following formula:

[0055]

[0056] wherein m 柱塞 is the oil supplement of the primary plunger pump B1 or the secondary plunger pump B2, in kg; m l is the leakage mass of the gap between the piston and the cylinder of the diaphragm compressor, in kg; X m is the design coefficient, greater than 1, preferably 3-5; p oil is the density of the hydraulic oil, in kg·m -3 ; d zs is the diameter of the plunger 5 of the plunger pump, in m; L zs is the stroke of the plunger 5 of the plunger pump, in m. Q l = m l *T, T = 60 / n, T is the reciprocating cycle time of the piston of the diaphragm compressor, in s, and n is the speed of the diaphragm compressor, in r / min.

[0057] The diameter and stroke of the plunger 5 of the plunger pump of the diaphragm compressor can be determined by the above method, thereby determining the basic design parameters of the primary plunger pump B1 and the secondary plunger pump B2.

[0058] In a possible implementation, the design and selection of the plunger pump spring 4 need to consider that the spring force is greater than the static friction force when the plunger 5 of the plunger pump is ready to move from the top dead center to the bottom dead center, and the spring force is always greater than the sliding friction force during the movement of the plunger 5 of the plunger pump from the top dead center to the bottom dead center. Since the difference between the static friction force and the sliding friction force is much smaller than the absolute amount of change of the spring force, only the plunger pump friction during the sliding process is considered for the selection of the plunger pump spring. The plunger pump friction is mainly composed of the sliding friction caused by the compression reaction force of the sealing ring and the viscous moving friction caused by the leaked hydraulic oil. The minimum spring force of the plunger pump spring 4 can be described as follows:

[0059]

[0060] wherein F spring,min is the minimum spring force of the plunger pump spring 4 at the bottom dead center position, in N; E is the elastic modulus of the sealing ring in the plunger sealing groove, in Pa; H is the wire diameter of the sealing ring in the plunger sealing groove, in m; h is the groove depth of the sealing ring in the plunger sealing groove, in m; D is the inner diameter of the sealing ring in the plunger sealing groove, in m; b is the sealing groove width in the plunger sealing groove, in m; m is the Poisson's ratio of the sealing ring in the plunger sealing groove; l is the characteristic length of the plunger 5 of the plunger pump in the axial direction contacting the leaked hydraulic oil, in m; v is the kinematic viscosity of the hydraulic oil, in mm 2 ·s -1 ; and p is the density of the hydraulic oil, in kg·m -3; u is the moving speed of the plunger pump plunger 5, in m·s -1 ; s is the circumferential gap of the plunger pump plunger 5, in m.

[0061] The minimum spring force required by the plunger pump spring 4 of the diaphragm compressor can be determined by the above method, so as to determine the specific parameters of the plunger pump spring 4 in cooperation with the overall structural size of the designed first-stage plunger pump B1 and second-stage plunger pump B2.

[0062] Please refer to Figure 4 In a possible implementation, there is a certain phase difference between the reciprocating direction of the diaphragm compressor piston and the reciprocating direction of the plunger pump plunger 5, and the phase difference is determined by the relative high pressure and the relative low pressure of the diaphragm compressor. The relative high pressure and the relative low pressure are calculated as follows:

[0063]

[0064] In the formula, p relative,high is the relative high pressure, in Pa; χ h is the relative high pressure coefficient, preferably 0.3-0.8; p oil,max is the maximum oil pressure, in Pa; χ l is the relative low pressure coefficient, preferably 0.2-0.6; p stage,1 is the first-stage instantaneous oil pressure of the diaphragm compressor, in Pa; p stage,2 is the second-stage instantaneous oil pressure of the diaphragm compressor, in Pa; t s is the phase angle at which the oil supplement of the overflow valve of the diaphragm compressor starts, t e is the phase angle at which the oil supplement of the overflow valve of the diaphragm compressor ends.

[0065] Further, the first-stage plunger pump B1 and the second-stage plunger pump B2 are symmetrically installed at 180°, and by selecting the relative high pressure coefficient χ h and the relative low pressure coefficient X l , the phase angle t s at which the oil supplement of the overflow valve of the diaphragm compressor starts and the phase angle t e at which the oil supplement of the overflow valve of the diaphragm compressor ends satisfy:

[0066]

[0067] The relative movement direction of the first-stage plunger pump B1 and the second-stage plunger pump B2 and the relative movement direction of the piston movement of the two-stage diaphragm compressor exist a phase angle which satisfies the following relationship:

[0068]

[0069] The best installation angle range of the first-stage piston pump B1 and the second-stage piston pump B2 of the double-stage reciprocating diaphragm compressor can be determined by the method, and the installation angle can minimize the working pressure of the piston pump and fully supply the oil.

[0070] Further, as shown in Figure 5 , the first stage of the double-stage reciprocating diaphragm compressor inhales air within a time range, and compresses the exhaust gas within a time range, and the best time for the piston pump to start supplying oil to the membrane head is within the time range [t s , t e ], and the best time for the piston pump to start supplying oil is within the time range [t stage , t oil,max ], and the design of the second-stage piston pump is the same.

[0071] Please refer to Figure 6 , another embodiment of the present application also provides a design method of a double-stage reciprocating piston pump structure diaphragm compressor as described, comprising the following steps:

[0072] Define the oil pressure p stage in the cylinder, and the dimensionless oil pressure is described as Here, the maximum oil pressure p oil,max is taken as the nominal exhaust gas pressure of the diaphragm compressor, and the nominal dimensionless oil pressure is described as The variation curve of the dimensionless oil pressure of the diaphragm compressor with the crank angle is obtained through simulation calculation based on the thermal process of AMESim, the relatively high pressure coefficient χ h = 0.6, and the relatively low pressure coefficient χ l = 0.2, and the intersection points of the first-stage dimensionless oil pressure curve are A1, B1, C1, and D1, and the intersection points of the second-stage dimensionless oil pressure curve are A2, B2, C2, and D2, and the following is calculated: According to the foregoing design method of the overflow valve, the best oil supply start phase angle t s of the first-stage and second-stage overflow valves is taken as 25.2°-79.2°, the oil supply end phase angle t e of the overflow valve is taken as 255.6°-298.8°, in order to satisfy , the best oil supply start phase angle t s is taken as 75.6°-79.2°, the relatively high pressure coefficient χ h = 0.38, and the relatively low pressure coefficient χ l = 0.3, and finally t s = 75° is taken, and the phase difference between the piston movement direction of the piston pump and the piston movement direction of the two-stage membrane head is determined as

[0073] The plunger pump structure of the embodiment of the application adopts the method of simplifying the hydraulic oil leakage gap channel model, quantitatively describes and calculates the leakage amount of the oil cavity of the diaphragm compressor, and quantitatively describes and calculates the stroke volume of the plunger pump cavity by the correction coefficient method to replace the method of determining the stroke volume of the plunger pump by the empirical ratio of the oil cavity volume; the minimum spring force required for the operation of the plunger pump is quantitatively described and calculated by considering the friction force of the sealing ring of the plunger pump and the viscous friction force of the hydraulic oil, and the selection of the plunger pump spring 4 is optimized; at the same time, the optimal installation phase difference of the plunger pump is quantitatively described and calculated by taking the minimum working pressure of the plunger pump as the optimization target, the optimal oil injection time range and the starting oil supplementing moment of the plunger pump are determined, which is beneficial to improve the service life of the plunger pump of the diaphragm compressor and the overall reliability of the diaphragm compressor. The diaphragm compressor plunger pump designed according to the method of the embodiment of the application has simple structure, small volume, flexible and convenient use, and is convenient for serial production or customization.

[0074] Although the application has been described above with reference to particular embodiments, it is to be understood that the skilled person can make many modifications to the configurations and details disclosed in the application within the principles and scope of the application. The scope of protection of the application is determined by the appended claims, and the claims are intended to cover all modifications included in the literal meaning or scope of the technical features of the claims.

Claims

1. A two-stage reciprocating-piston pump structure diaphragm compressor, characterized by: The compressor end cover (A), the first-stage plunger pump (B1), the second-stage plunger pump (B2), the plunger pump connecting bolt (C), the first-stage eccentric structure (D1) and the second-stage eccentric structure (D2) are included; the first-stage plunger pump (B1) and the second-stage plunger pump (B2) are respectively installed on the symmetrical two sides of the compressor end cover (A) through the plunger pump connecting bolt (C) to realize power balance; the plunger pump body (1) of the first-stage plunger pump (B1) and the second-stage plunger pump (B2) and the compressor end cover (A), the plunger pump body (1) and the plunger pump bushing (3), the plunger pump body (1) and the plunger pump cover (2) and the plunger pump bushing (3) and the internal plunger pump plunger (5) are all sealed by the sealing ring to realize hydraulic oil sealing; the first-stage plunger pump (B1) and the second-stage plunger pump (B2) are respectively driven by the first-stage eccentric structure (D1) and the second-stage eccentric structure (D2) and the counteracting force of the plunger pump spring (4) and the friction force of the plunger pump plunger (5) to realize the processes of oil feeding, pressure increasing, oil discharging and pressure decreasing of the plunger pump. The reciprocating movement direction of the diaphragm compressor piston and the reciprocating movement direction of the plunger pump plunger (5) have a phase difference, and the phase difference is determined by the relative high pressure and the relative low pressure of the diaphragm compressor; the relative high pressure and the relative low pressure of the diaphragm compressor are calculated according to the following formula: wherein, is a relatively high pressure; is a relatively high pressure coefficient; is a maximum oil pressure; is a relatively low pressure coefficient; is a diaphragm compressor first stage instantaneous oil pressure; is a diaphragm compressor second stage instantaneous oil pressure; is a diaphragm compressor relief valve oil replenishment start phase angle, is a diaphragm compressor relief valve oil replenishment end phase angle; The primary plunger pump (B1) and the secondary plunger pump (B2) are symmetrically installed at 180°, and by selecting a relatively high pressure coefficient and a relatively low pressure coefficient , the phase angle at which the diaphragm compressor overflow valve starts to supplement oil and the phase angle at which the diaphragm compressor overflow valve stops supplementing oil satisfy: The relative movement direction of the primary plunger pump (B1) and the secondary plunger pump (B2) and the relative movement direction of the piston movement of the double-stage reciprocating diaphragm compressor exist a phase angle , which meets the following relationship: The first stage of the diaphragm compressor sucks in air during the time interval compresses the exhaust air during the time interval The best moment for the plunger pump to start refilling the membrane head is at the moment The best moment for the plunger pump to start refilling the membrane head is at the moment .

2. The dual stage, double acting, reciprocating piston pump structure, membrane compressor according to claim 1, characterized in that: The plunger pump body (1) of the first-stage plunger pump (B1) and the second-stage plunger pump (B2) and the compressor end cover (A) are sealed by the sealing ring in the fourth sealing groove (14) of the plunger pump body to realize hydraulic oil sealing, the plunger pump body (1) and the plunger pump bushing (3) are sealed by the sealing ring in the second sealing groove (12) and the third sealing groove (13) of the plunger pump body to realize hydraulic oil sealing, the plunger pump cover (2) and the plunger pump body (1) are sealed by the sealing ring in the first sealing groove (11) of the plunger pump body to realize hydraulic oil sealing, and the plunger pump bushing (3) and the plunger pump plunger (5) are sealed by the sealing ring in the plurality of plunger sealing grooves (51) to realize hydraulic oil sealing.

3. The dual stage, double acting, reciprocating piston pump structure, membrane compressor according to claim 2, characterized in that: The first-stage plunger pump (B1) and the second-stage plunger pump (B2) are respectively driven by the first-stage eccentric structure (D1) and the second-stage eccentric structure (D2), and the plunger pump plunger (5) moves to the compression direction of the plunger pump spring (4) in this process, the volume of the plunger pump oil cavity decreases, the processes of pressure increasing and oil discharging of the hydraulic oil in the plunger pump oil cavity are realized, and the plunger pump spring (4) provides the counteracting force against the driving force of the eccentric and the friction force of the plunger pump plunger (5); when the plunger pump plunger (5) moves to the top dead center position, the plunger pump plunger (5) is driven by the plunger pump spring (4) to move to the bottom dead center position against the counteracting force of the eccentric and the friction force of the plunger pump plunger (5), and the plunger pump plunger (5) is driven to move to the bottom dead center position; in this process, the volume of the plunger pump oil cavity increases, the processes of pressure decreasing and oil feeding of the hydraulic oil in the plunger pump oil cavity are realized, and the plunger pump plunger (5) reaches the bottom dead center position.

4. The dual stage, double acting, reciprocating piston pump structure, membrane compressor according to claim 2, wherein: The volume chamber volume of the primary plunger pump (B1) and the secondary plunger pump (B2) is greater than the leakage passage gap L of the diaphragm compressor piston xl ; Membrane compressor piston leakage passage gap L xl The leakage amount of the gap L is calculated by the following equation: In the formula, For the piston leakage passage clearance L of the diaphragm compressor xl Leakage flow; The diameter of the piston in the diaphragm compressor; For the piston leakage passage clearance L of the diaphragm compressor xl The gap; The viscosity of the hydraulic oil; For the piston leakage passage clearance L of the diaphragm compressor xl Length; For the piston leakage passage clearance L of the diaphragm compressor xl The pressure difference on both sides; This represents the maximum oil pressure of the diaphragm compressor. Atmospheric pressure; This refers to the number of sealing rings on the piston of the diaphragm compressor.

5. The dual stage, double acting, reciprocating piston pump structure, membrane compressor according to claim 4, characterized in that: The volume cavity volume of the first-stage plunger pump (B1) and the second-stage plunger pump (B2) is calculated according to the following formula: wherein is the oil makeup for a single piston pump (B1) or a two-stage piston pump (B2); is the piston leakage passage gap leakage mass for a diaphragm compressor; is the design coefficient, greater than 1; is the hydraulic oil density; is the diameter of the piston pump piston (5); is the stroke of the piston pump piston (5).

6. The dual stage, double acting, reciprocating piston pump structure, membrane compressor according to claim 1, characterized in that: The counteracting force of the plunger pump spring (4) is calculated according to the following formula: wherein Fminis the spring force of the minimum plunger pump spring (4) when the bottom dead center position; Eis the modulus of elasticity of the sealing ring in the plunger sealing groove; Lis the linear diameter of the sealing ring in the plunger sealing groove; Dsis the groove depth of the sealing ring in the plunger sealing groove; Dis the inner diameter of the sealing ring in the plunger sealing groove; Wsis the sealing groove width in the plunger sealing groove; Psis the Poisson's ratio of the sealing ring in the plunger sealing groove; Lpumpis the characteristic length of the plunger pump plunger (5) in the axial direction contacting the leaked hydraulic oil; Vis the kinematic viscosity of the hydraulic oil; pis the density of the hydraulic oil; Vpumpis the moving speed of the plunger pump plunger (5); Cpumpis the circumferential gap of the plunger pump plunger (5).

7. A method of designing a two-stage reciprocating-piston pump-structure diaphragm compressor as claimed in any one of claims 1 to 6, characterized in that, The steps include: The dimensionless oil pressure is defined as The dimensionless oil pressure is defined as The maximum oil pressure is taken here The nominal discharge pressure is designed for the diaphragm compressor, and the nominal dimensionless oil pressure is described as =1, the curve of the dimensionless oil pressure of the diaphragm compressor with the crank angle is obtained by the simulation calculation of the thermodynamic process. By selecting the relatively high pressure coefficient and the relatively low pressure coefficient , the starting phase angle and the ending phase angle of the diaphragm compressor overflow valve oil supplement meet: Determine the phase angle of the relative movement direction of the primary piston pump (B1) and the secondary piston pump (B2) and the relative movement direction of the piston movement of the double-stage piston compressor Determine the installation angle range of the primary piston pump (B1) and the secondary piston pump (B2).

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