US20110256001A1 - Reciprocating Compressor - Google Patents

Reciprocating Compressor Download PDF

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Publication number
US20110256001A1
US20110256001A1 US13/022,652 US201113022652A US2011256001A1 US 20110256001 A1 US20110256001 A1 US 20110256001A1 US 201113022652 A US201113022652 A US 201113022652A US 2011256001 A1 US2011256001 A1 US 2011256001A1
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United States
Prior art keywords
pressure
piston
low
pressure piston
cylinder
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Abandoned
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US13/022,652
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English (en)
Inventor
Hisatoshi Kobayashi
Akito OHATA
Susumu Sakamoto
Koichi Tashiro
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Hitachi Industrial Equipment Systems Co Ltd
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Hitachi Industrial Equipment Systems Co Ltd
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Assigned to HITACHI INDUSTRIAL EQUIPMENT SYSTEMS CO., LTD. reassignment HITACHI INDUSTRIAL EQUIPMENT SYSTEMS CO., LTD. ASSIGNMENT OF ASSIGNORS INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: SAKAMOTO, SUSUMU, KOBAYASHI, HISATOSHI, OHATA, AKITO, TASHIRO, KOICHI
Publication of US20110256001A1 publication Critical patent/US20110256001A1/en
Abandoned legal-status Critical Current

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B25/00—Multi-stage pumps
    • F04B25/005—Multi-stage pumps with two cylinders
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B39/00—Component parts, details, or accessories, of pumps or pumping systems specially adapted for elastic fluids, not otherwise provided for in, or of interest apart from, groups F04B25/00 - F04B37/00
    • F04B39/0005—Component parts, details, or accessories, of pumps or pumping systems specially adapted for elastic fluids, not otherwise provided for in, or of interest apart from, groups F04B25/00 - F04B37/00 adaptations of pistons

Definitions

  • the present invention relates to a reciprocating compressor in which pistons reciprocate within cylinders.
  • a piston with a piston body fixed to a connecting rod is used in the low-pressure compressor unit, so that the piston reciprocates while oscillating (tilting) within a cylinder.
  • a piston with a piston body oscillating with respect to a leading end of a connecting rod is used in the high-pressure compressor unit, so that a leading end of the piston reciprocates without tilting within a cylinder.
  • the piston is formed without consideration of a tilt angle of the piston within the low-pressure or high-pressure cylinder. It is therefore not possible to improve the sealing performance and lifetime of the piston.
  • an object of the present invention is to provide a reciprocating compressor with pistons of low- and high-pressure compressor units being compact and having high sealing performance by forming the pistons in consideration of tilt angles of the pistons within cylinders.
  • a reciprocating compressor including: a low-pressure compressor unit having a low-pressure piston and a low-pressure cylinder for compressing air with the low-pressure piston reciprocating while oscillating within the low-pressure cylinder; a high-pressure compressor unit having a high-pressure piston and a high-pressure cylinder for further compressing the air compressed in the low-pressure compressor unit with the high-pressure piston reciprocating while oscillating within the high-pressure cylinder; and a motor for driving the low-pressure compressor unit and the high-pressure compressor unit.
  • a maximum tilt angle during oscillation of the low-pressure piston is made larger than a maximum tilt angle during oscillation of the high-pressure piston.
  • a reciprocating compressor including: a motor having a rotating shaft; a low-pressure compressor unit having a low-pressure cylinder and a low-pressure piston for compressing air; and a high-pressure compressor unit having a high-pressure cylinder and a high-pressure piston for further compressing the air compressed in the low-pressure compressor unit.
  • the low-pressure piston and the high-pressure piston each include: an eccentric portion for performing eccentric motion with rotation of the rotating shaft of the motor; a connecting rod extending from the eccentric portion; and a piston body provided on a leading end of the connecting rod.
  • FIG. 1 is a diagram showing a low-pressure compressor unit and a high-pressure compressor unit of a compressor according to an embodiment of the present invention
  • FIG. 2 is an enlarged view of the high-pressure compressor unit according to the embodiment of the present invention.
  • FIG. 3 is a general view of the reciprocating compressor according to the embodiment of the present invention.
  • FIG. 4 is a diagram showing oscillating motion of a piston according to the embodiment of the present invention.
  • FIGS. 1 to 4 An embodiment of the present invention will be described with reference to FIGS. 1 to 4 .
  • the compressor according to this embodiment has a crankcase 1 .
  • a motor 3 having a shaft (a rotating shaft) 2 is attached to the crankcase 1 .
  • a low-pressure piston 4 including a connecting rod 4 A and a piston body 5 , and a high-pressure piston 8 including a connecting rod 8 A and a piston body 11 are attached to the shaft 2 of the motor 3 through eccentric portions 7 and 13 for converting a rotational motion to a reciprocating motion.
  • the low-pressure piston body 5 is composed of a retainer 5 A
  • the high-pressure piston body 11 is composed of a base 11 A and a top 12 A.
  • the low-pressure piston body 5 is provided with a lip ring 6 .
  • the high-pressure piston body 11 is provided with a lip ring 9 and a piston ring 10 .
  • the lip ring 6 is provided with a skirt portion (a lip portion) for increasing a contact area with a cylinder 14 .
  • the skirt portion is mounted facing a low-pressure compression chamber, thereby allowing the ensuring of the sealing performance between the piston 4 and the cylinder 14 .
  • the lip ring 6 , the lip ring 9 , and the piston ring 10 are made of resin material having excellent resistance to abrasion and self-lubricating property and each formed in an approximately annular shape.
  • the piston ring 10 is substantially rectangular in cross section, and has a uniform radial width over the almost entire circumference thereof. Also, an abutment joint (not shown) is circumferentially formed on the piston ring 10 . This abutment joint allows expansion and contraction in diameter while maintaining the sealing performance.
  • the inner diameter of the piston ring 10 in a state brought into contact with an inner peripheral surface of a high-pressure cylinder 17 to be described later is larger than the smallest diameter of a piston ring groove into which the piston ring 10 is mounted.
  • the piston ring 10 is radially movable with respect to the high-pressure piston 8 .
  • the lip ring 9 is mounted in the direction opposite to the lip ring 6 (the skirt portion of the lip ring 9 faces the crankcase), so that the centers of the lip ring 9 and the piston body 11 are aligned. Also, when the cylinder 17 is assembled in the crankcase 1 , the lip ring 9 is brought into contact with an inner wall surface of the cylinder 17 to define an assembly position of the cylinder 17 , so that the centers of the cylinder 17 and the piston body 11 are aligned. Thus, centering of the piston ring 10 mounted on the piston body 11 , and the cylinder 17 can be performed. Further, the lip ring 9 can prevent contact between the piston body 11 and the cylinder 17 when the piston ring 10 becomes worn, and therefore the lifetime of the piston body 11 and the cylinder 17 can be improved.
  • the lip ring 9 is interposed between the base 11 A and the connecting rod 8 A.
  • compression heat generated in the cylinder 17 can be prevented from being transmitted from the piston body 11 to the connecting rod 8 A, so that the temperature of a large end can be lowered. It is therefore possible to improve the lives of bearings provided on outer peripheries of the eccentric portions 7 and 13 .
  • the lip ring 6 is mounted on the low-pressure piston 4 , thereby allowing reduction in manufacturing cost. Also, the lip ring 9 and the piston ring 10 are mounted on the high-pressure piston 8 , thereby allowing improvement in assembly performance, sealing performance, and resistance to abrasion.
  • the crankcase 1 is mounted with the low-pressure cylinder 14 , an air valve 15 , a cylinder head 16 , the high-pressure cylinder 17 , an air valve 18 , and a cylinder head 19 .
  • the low-pressure piston 4 including the low-pressure connecting rod 4 A, and the high-pressure piston 8 including the high-pressure connecting rod 8 A compression chambers are formed within the respective cylinders.
  • the eccentric portions 7 and 13 allow the low- and high-pressure pistons 4 and 8 to reciprocate within the cylinders 14 and 17 , respectively, thereby driving the low-pressure compressor unit including the low-pressure piston 4 and cylinder 14 , and the high-pressure compressor unit including the high-pressure piston 8 and cylinder 17 .
  • the compressor according to the embodiment of the present invention has a two-stage compression structure in which the low-pressure compressor unit sucks air into the cylinder 14 through the cylinder head 16 and the air valve 15 from the atmospheric pressure and compresses the sucked air to discharge the compressed air to the high-pressure compressor unit through a pipe (not shown), and then the high-pressure compressor unit sucks the compressed air discharged from the low-pressure compressor unit and further compresses the air to discharge the air to a storage tank.
  • the low- and high-pressure compressor units and the pipe is air-cooled by the rotation of a fan 20 mounted coaxially with the shaft 2 .
  • the above-described problems are avoided by adopting the following two-stage compression structure.
  • the two-stage compression structure in view of low vibration, low noise, miniaturization (downsizing of a motor, weight reduction of a compressor unit).
  • the respective pressure ratios on the low- and high-pressure sides are reduced, relative to the case of the single-stage compression, thereby allowing improvement in efficiency, reduction in heat generation, and reduction in performance deterioration caused by the heat generation described above. Furthermore, torque fluctuations of the motor can be reduced and therefore low vibration and low noise can be realized.
  • the two-stage compressor disclosed in Japanese Patent Application Laid-Open Publication No. H11-62822 adopts a reciprocating piston structure in which a connecting rod is connected to a piston through a needle bearing on the high-pressure side.
  • a needle bearing is removed, and a rocking piston mechanism with the connecting rod and the piston body integrally formed is employed, thereby realizing durability improvement by reducing movable portions, weight reduction and low noise, and reduction in cost by reducing the number of components.
  • the oscillating compressor disclosed in Japanese Patent Application Laid-Open Publication No. 2007-32532 includes the pistons formed without relatively considering the angles at which the pistons are tilted within the low- and high-pressure cylinders. Consequently, in the low-pressure compressor unit, the piston tilt angle is minimized, in spite of the fact that, even when the piston tilt angle is made larger than that of the high-pressure side, its influence on the sealing performance and abrasion property is small. For this reason, the downsizing and weight reduction of the piston cannot be sufficiently achieved. Also, in the high-pressure compressor unit, the piston tilt angle is not designed to be made sufficiently small, in spite of the fact that, when the piston tilt angle is made too large, its influence on the sealing performance and abrasion property increases. For this reason, the sealing performance and lifetime of the piston cannot be achieved.
  • the low-pressure piston 4 and the high-pressure piston 8 include the rocking piston mechanism in which: the connecting rods 4 A and 8 A and the piston bodies 5 and 11 are integrally formed; the piston bodies 5 and 11 are tilted as the connecting rods 4 A and 8 A are tilted; and the piston bodies 5 and 11 reciprocate while oscillating within the cylinders 14 and 17 .
  • the compressor according to this embodiment is designed in consideration of the angles at which the low-pressure piston 4 and the high-pressure piston 8 are tilted within the cylinders 14 and 17 , respectively.
  • the motor powers that is, the shaft powers (the workloads w) of the low- and high-pressure sides increase with an increase in the respective pressure ratios of suction pressure to discharge pressure.
  • the shaft power of the two-stage compressor depends on the sum of the shaft powers of the low- and high-pressure sides. Also, the shaft power decreases as the pressure ratios on the low- and high-pressure sides decrease. That is to say, when the pressure ratios on the low- and high-pressure sides become equal, the shaft power becomes minimum.
  • various sizes (such as bore diameter and stroke) of the low- and high-pressure pistons 4 and 8 are designed based on the shaft power in consideration of the maximum pressure of the compressor.
  • Lad represents the theoretical adiabatic aerodynamic force
  • Qs represents the actual air volume in suction condition
  • Ps represents the suction absolute pressure
  • Pd represents the discharge absolute pressure
  • ⁇ represents the specific heat ratio
  • D represents the bore diameter
  • S represents the stroke
  • N represents the rotational speed
  • ⁇ v represents the volumetric efficiency
  • the motor power decreases as the pressure ratios on the low- and high-pressure sides in the two-stage compression and the difference in shaft power therebetween decrease.
  • the various sizes of the low- and high-pressure pistons 4 and 8 are calculated as expressed by the following equations (4) and (5):
  • L represents the total shaft power
  • Ls 1 represents the low-pressure required shaft power
  • Ls 2 represents the high-pressure required shaft power
  • Pm represents the intermediate absolute pressure
  • P 1 represents the low-pressure suction absolute pressure
  • P 2 represents the high-pressure discharge absolute pressure
  • Ls 1 and Ls 2 are expressed as follows:
  • Qs 1 represents the actual air volume in suction condition on the low-pressure side.
  • Qs 2 represents the actual air volume in suction condition on the high-pressure side.
  • D 1 represents the low-pressure bore diameter
  • D 2 represents the high-pressure bore diameter
  • S 1 represents the low-pressure stroke
  • S 2 represents the high-pressure stroke
  • respective bore diameters and strokes of the low-pressure piston 4 and the high-pressure piston 8 are determined by the equations (12) and (13).
  • P 2 must be made sufficiently large relative to P 1 to obtain compressed air at high pressure. That is to say, in the equation (12), Pm must be made sufficiently large relative to P 1 .
  • D 1 must be made sufficiently large relative to D 2
  • S 1 must be made sufficiently large relative to S 2 to satisfy the equation (12) (to at least approximate the left- and right-hand values).
  • the low-pressure bore diameter D 1 and the high-pressure bore diameter D 2 are preferably set to satisfy D 1 >D 2 . This is because, in the case of the two-stage compression, the pressure in the high-pressure compression chamber is higher than that in the low-pressure compression chamber, and therefore the area of the high-pressure piston 8 is reduced to reduce a load thereon so as to reduce a load F of the piston 8 applied in the longitudinal direction of the connecting rod 8 A from a top surface of the piston 8 and downsize the bearing provided on the outer periphery of the eccentric portion 13 on the center side of the connecting rod 8 A.
  • the center-of-gravity balance of the product is important, especially in portable air compressors.
  • the portable air compressor is mounted with a compressor body 21 including the low- and high-pressure compressor units described in FIG. 1 , and the motor 3 , for example, on a pair of air tanks 22 (roughly in the center thereof).
  • various auxiliary components in particular, a reducing valve 23 ( 26 ), a pressure gauge 24 ( 27 ), and an air outlet coupler 25 ( 28 ), which are large in mass, are mounted symmetrically with respect to the compressor body 21 , thereby performing layout in consideration of the center-of-gravity balance of the product.
  • the center-of-gravity balance of the compressor body itself is also important.
  • the low- and high-pressure bore diameters D 1 and D 2 and the low- and high-pressure strokes S 1 and S 2 are determined so that the low- and high-pressure sides are made equal in shaft power and pressure ratio.
  • the two-stage compression structure according to the embodiment of the present invention as shown in FIG. 1 includes the opposed two cylinders, and therefore the center-of-gravity balance of the compressor body 21 is greatly influenced by the lengths and sizes of the low- and high-pressure cylinders 14 and 17 , air valves 15 and 18 , cylinder heads 16 and 19 protruding from the crankcase 1 .
  • the air cooled by the cooling fan 20 mounted on the compressor body 21 is less likely to reach the larger one in the length from the compressor body 21 to the cylinder head 16 ( 19 ), thereby producing an increase in temperature.
  • performance deterioration and lifetime shortening might be caused.
  • the compressor body 21 is mounted on the air tanks 22 , the cylinder head of the larger one in length protrudes from between the air tanks 22 . As a result, a disadvantage of an increase in dimension of the product might occur.
  • the compressor body 21 is mounted in such a manner that the axial direction of the shaft 2 and the longitudinal direction of the air tanks 22 are perpendicular to each other, thereby satisfying both miniaturization and weight balance.
  • the present invention is not limited thereto, and the compressor body 21 may be mounted in such a manner that the axial direction of the shaft 2 and the longitudinal direction of the air tanks 22 face in the same direction. In this case, by locating the shaft 2 in between the two air tanks 22 , the center-of-gravity balance can be ensured. Furthermore, the size is determined so that the cylinder head is prevented from protruding from between the air tanks 22 , thereby allowing realization of miniaturization.
  • D 1 , D 2 , S 1 , and S 2 are determined in consideration of the foregoing.
  • D 1 is preferably set twice or less as large as D 2 .
  • S 1 ⁇ S 2 is set, Pm becomes four times or less as large as P 1 , and thus compressed air at sufficiently high pressure cannot be obtained. Therefore, in order to obtain compressed air at sufficiently high pressure, S 1 must be designed large relative to S 2 , and S 1 >S 2 must be set.
  • the connecting rod 4 A ( 8 A) becomes oblique with respect to a central shaft 20 of the cylinder 14 ( 17 ) by the eccentricity of the eccentric portion 7 ( 13 ).
  • the maximum tilt angle of the piston 4 ( 8 ) with respect to the cylinder 14 ( 17 ) during oscillation will be described.
  • the tilt angle ⁇ is determined by, the length l of the connecting rod 4 A ( 8 A), that is, the length from the center of the eccentric portion 7 ( 13 ) to the leading end (top surface) of the piston body 5 ( 11 ), and the eccentric amount r of the eccentric portion 7 ( 13 ) with respect to the shaft 2 of the motor 3 , as expressed by the following equation (14):
  • the tilt angle ⁇ is minimized while considering differences in the bore diameter D and the connecting rod length l between the low- and high-pressure sides, and the low- and high-pressure bore diameters D and strokes S (the stroke S equals two multiplied by the eccentric amount r), and the connecting rod lengths l are determined by the foregoing equations while preventing the low-pressure maximum tilt angle ⁇ 1 from becoming smaller than the high-pressure maximum tilt angle ⁇ 2 .
  • the performance deterioration due to leakage of the compressed air is avoided, in particular, the performance deterioration on the high-pressure side can be prevented.
  • compressed air at high pressure can be obtained while preventing the deterioration in the center-of-gravity balance of the product and unequal cooling.
  • the piston body 5 of the low-pressure piston 4 is mounted with the lip ring 6 that is flexible and has high followability to a change in the gap between the cylinder 14 and the lip ring 6 , thereby reducing the tendency to cause the performance deterioration due to leakage of compressed air from a gap formed in an oscillating direction of the piston 4 .
  • the piston body 11 of the high-pressure piston 8 is mounted with the piston ring 10 requiring stiffness under high pressure and temperature.
  • the piston ring 10 is poor in followability to a change in the gap relative to the lip ring 6 , and thus the performance deterioration due to leakage of compressed air is likely to be caused. It is therefore necessary to prevent the high-pressure side from being affected by leakage of compressed air, in consideration of the angle at which the piston 8 is tilted with respect to the cylinder 17 .
  • the connecting rod 4 A ( 8 A) moves toward the top dead center position and the bottom dead center position during the suction and discharge processes, the connecting rod 4 A ( 8 A) becomes oblique with respect to the central shaft 20 of the cylinder 14 ( 17 ) by the eccentricity of the eccentric portion 7 ( 13 ).
  • the contact surface between the lip ring 6 (the piston ring 10 ) and the cylinder 14 ( 17 ) is formed in an elliptical shape (as viewed from the upper side of the cylinder central shaft) with the oscillating direction (the horizontal direction in FIG. 4 ) corresponding to the long axis.
  • a gap between a side 6 A ( 9 A) in the oscillating direction of the lip ring 6 (the piston ring 10 ) and the cylinder 14 ( 17 ) is likely to be formed.
  • the performance deterioration is likely to be caused due to leakage of compressed air from the gap.
  • the compressor is designed so as to prevent the high-pressure maximum tilt angle ⁇ 2 from becoming larger than the low-pressure maximum tilt angle ⁇ 1 . Therefore, a maximum gap T 2 formed between the high-pressure piston 8 and the cylinder 17 can be prevented from becoming larger than a maximum gap T 1 formed between the low-pressure piston 4 and the cylinder 14 . This allows prevention of occurrence of the performance deterioration due to leakage of compressed air, in particular, on the high-pressure compressor unit.
  • the piston ring 10 having stiffness higher than the lip ring is mounted on the high-pressure piston body 11 , thereby also reducing the tendency to the performance deterioration due to abrasion.
  • the embodiment of the present invention has been described by using the case where the lip ring 6 is provided on the low-pressure piston body 5 , a piston ring may be provided in place of the lip ring, in the same manner as the high pressure side.
  • followability to the gap is reduced on the low-pressure compressor unit, however, since pressure in the compression chamber of the low-pressure compressor unit is lower than that of the high-pressure compressor unit, the low-pressure piston body 5 with the piston ring may be adopted by taking a measure such as reducing the thickness of the piston ring to thereby increase the followability.
  • the low-pressure piston body 5 is provided with the piston ring, the performance deterioration due to abrasion of the piston 4 can be also prevented on the low-pressure compressor unit.
  • the gas load F to be applied toward the connecting rod central axis from the piston top surface in the case where the piston body 5 ( 11 ) is the upper side and the connecting rod 4 A ( 8 A) is the lower side
  • a horizontal component f of the gas load F occur, thereby pressing the lip ring 6 ( 9 ) against the cylinder 14 ( 17 ).
  • the surface abrasion of the lip ring 6 ( 9 ) or the cylinder 14 ( 17 ) advances, which can result in performance deterioration.
  • the gas load F is large especially on the high-pressure compressor unit, it is necessary to reduce the surface abrasion of the lip ring 9 , the piston ring 10 , and the cylinder 17 on the high-pressure compressor unit to prevent the performance deterioration.
  • the horizontal component f of the gas load F increases as the tilt angle ⁇ corresponding to the angle at which the piston 4 ( 8 ) is tilted with respect to the cylinder 14 ( 17 ) increases.
  • the compressor is designed so as to prevent the high-pressure maximum tilt angle ⁇ 2 from becoming larger than the low-pressure maximum tilt angle ⁇ 1 . Therefore, especially on the high-pressure compressor unit in which the surface abrasions of the lip ring 9 , the piston ring 10 , and the cylinder 17 could be a problem, it is possible to reduce the abrasions thereof and prevent the performance deterioration.
  • the low-pressure piston 4 and the high-pressure piston 5 are constructed according to the above-described dimensional relationship therebetween, thereby minimizing the tilt angle ⁇ .
  • the low- and high-pressure bore diameters D and strokes S (the stroke S equals two multiplied by the eccentric amount r), and connecting rod lengths l are determined by the foregoing equations (1) to (14) so as to prevent the high-pressure maximum tilt angle ⁇ 2 from becoming larger than the low-pressure maximum tilt angle ⁇ 1 , thereby allowing prevention of the performance deterioration due to leakage of compressed air, especially on the high-pressure side.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Compressor (AREA)
  • Compressors, Vaccum Pumps And Other Relevant Systems (AREA)
US13/022,652 2010-04-14 2011-02-08 Reciprocating Compressor Abandoned US20110256001A1 (en)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
JP2010-092732 2010-04-14
JP2010092732A JP5380353B2 (ja) 2010-04-14 2010-04-14 往復動圧縮機

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US20130189132A1 (en) * 2012-01-23 2013-07-25 Shinano Kenshi Co., Ltd. Compressor and vacuum machine
US20130224051A1 (en) * 2012-02-27 2013-08-29 Shinano Kenshi Co., Ltd. Compressor and vacuum machine
US20140294636A1 (en) * 2013-03-29 2014-10-02 Hitachi Automotive Systems, Ltd. Reciprocating compressor
WO2016069667A1 (en) * 2014-10-29 2016-05-06 Emerson Climate Technologies, Inc. Reciprocating compressor system
US9447725B2 (en) 2012-03-23 2016-09-20 Shinano Kenshi Co., Ltd. Compressor and vacuum machine
US20170043645A1 (en) * 2015-08-11 2017-02-16 Ford Global Technologies, Llc Dynamically controlled vehicle cooling and heating system operable in multi-compression cycles
CN109882375A (zh) * 2019-04-10 2019-06-14 浙江北上新能源科技股份有限公司 一种两级压缩机
CN112696351A (zh) * 2021-01-07 2021-04-23 西南石油大学 一种双级压差式钻井泵活塞总成
US20230122679A1 (en) * 2021-10-14 2023-04-20 Beijingwest Industries Co., Ltd. Integrated air supply unit
CN117927448A (zh) * 2024-03-21 2024-04-26 苏州瑞玛精密工业股份有限公司 一种空压机、集成式气体供给装置及供气方法
US20250027488A1 (en) * 2023-07-21 2025-01-23 Weidong Lu Air compressor

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