CN114320915A - Pump body subassembly, compressor and refrigeration equipment of heating - Google Patents
Pump body subassembly, compressor and refrigeration equipment of heating Download PDFInfo
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- CN114320915A CN114320915A CN202111620461.5A CN202111620461A CN114320915A CN 114320915 A CN114320915 A CN 114320915A CN 202111620461 A CN202111620461 A CN 202111620461A CN 114320915 A CN114320915 A CN 114320915A
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Abstract
The application provides a pump body subassembly, compressor and refrigeration equipment of heating. The pump body subassembly includes: the air cylinder is internally provided with a compression cavity; the crankshaft comprises a main shaft, an eccentric shaft and an auxiliary shaft which are connected in sequence; the piston is arranged on the eccentric shaft and revolves in the compression cavity; the sliding sheet is abutted against the piston and reciprocates in the compression cavity; a main bearing; and, a secondary bearing; one end of one or two of the main bearing and the auxiliary bearing close to the cylinder is provided with a circular groove, and the central shaft of at least one circular groove is eccentrically arranged relative to the rotation center of the crankshaft. The application provides a pump body subassembly, with the eccentric settings of the centre of rotation of the relative bent axle of at least one ring groove, can increase the intensity of the inside wall subregion of ring groove to reduce the bent axle to this subregion skew, and then reduce the skew of piston to this subregion, so that the clearance is less between piston and compression cavity inner surface, and then reduces high-pressure gas to low-pressure gas's leakage, promotes compression efficiency.
Description
Technical Field
The application belongs to the technical field of compressors, and particularly relates to a pump body assembly, a compressor and refrigeration and heating equipment.
Background
The pump body assembly of rotary compressor is generally equipped with circular ring grooves on the main bearing and auxiliary bearing near the cylinder to reduce the abrasion between crankshaft and main bearing and auxiliary bearing. However, when the piston revolves in the compression chamber to compress gas, the pressure borne by the piston changes gradually, so that the crankshaft shakes around the rotation center when rotating, the inner side wall of the circular groove deforms, accordingly, the gap between the piston and the inner surface of the compression chamber is increased, high-pressure gas leaks to low-pressure gas, the energy efficiency of the compressor is reduced, and the compression efficiency is reduced.
Disclosure of Invention
An object of the embodiment of the application is to provide a pump body subassembly, compressor and refrigeration equipment of heating to when the crankshaft rotates among the pump body subassembly that exists among the solution prior art, the inside wall in ring groove can take place to warp, and makes the clearance increase between piston and the compression cavity internal surface, and leads to high pressure gas to experience the problem of leaking low pressure gas.
In order to achieve the above purpose, the embodiment of the present application adopts the following technical solutions: there is provided a pump body assembly comprising:
the air cylinder is internally provided with a compression cavity;
the crankshaft comprises a main shaft, an eccentric shaft and an auxiliary shaft which are connected in sequence;
the piston is arranged on the eccentric shaft and revolves in the compression cavity;
the sliding sheet is abutted against the piston and reciprocates in the compression cavity;
the main bearing is arranged at one end of the air cylinder, and a main sliding hole which is matched and sleeved on the main shaft is arranged in the main bearing; and the number of the first and second groups,
the auxiliary bearing is arranged at the other end of the cylinder, and an auxiliary sliding hole which is matched and sleeved on the auxiliary shaft is formed in the auxiliary bearing;
one end of one or two of the main bearing and the auxiliary bearing, which is close to the air cylinder, is provided with a circular groove, and the central shaft of at least one circular groove is eccentrically arranged relative to the rotation center of the crankshaft.
In an alternative embodiment, the central axis of the circular groove is relative to the rotation center of the crankshaft with reference to the central line of the reciprocating motion of the sliding vane, and the eccentric angle θ 1 along the revolution direction of the piston ranges from 0 degree to 90 degrees.
In an alternative embodiment, the center of the compression chamber is eccentrically disposed with respect to the rotational center of the crankshaft, and a minimum clearance position is formed at which the clearance between the inner surface of the compression chamber and the outer circumferential surface of the piston is at a minimum.
In an alternative embodiment, the direction from the central axis of the circular groove to the rotation center of the crankshaft is opposite to the direction from the minimum clearance position to the rotation center of the crankshaft, and the included angle along the revolution direction of the piston ranges from-50 degrees to 40 degrees.
In an alternative embodiment, the central axis of the circular groove is located on an extension of the minimum clearance position to the center of rotation of the crankshaft.
In an alternative embodiment, the included angle between the rotation center of the crankshaft and the position of the minimum clearance is 180 to 270 degrees in the revolving direction of the piston, based on the center line of the reciprocating motion of the sliding vane.
In an alternative embodiment, the rotation center of the crankshaft is up to the minimum clearance position with respect to the center line of the reciprocating motion of the sliding vane, and the included angle along the revolution direction of the piston is 230 degrees.
In an alternative embodiment, the main bearing and the secondary bearing are provided with the annular grooves.
In an alternative embodiment, the central axes of the circular grooves of the main bearing and the auxiliary bearing are eccentrically arranged relative to the revolution center of the crankshaft.
In an alternative embodiment, the two circular ring grooves have the same eccentric distance with respect to the rotation center of the crankshaft.
In an alternative embodiment, the inner side wall of the circular groove forms a circular bearing; in the axial direction of the compression chamber: the distance from the end face of the circular ring bearing to the cylinder is larger than the distance from the end face of the outer side wall of the circular ring groove to the cylinder.
In an alternative embodiment, the number of the cylinders is multiple, a partition plate is arranged between every two adjacent cylinders, the pistons are respectively arranged in the cylinders, and the eccentric shafts correspond to the pistons one by one.
It is another object of an embodiment of the present application to provide a compressor, which includes a housing, a motor installed in the housing, and the pump body assembly according to any one of the above embodiments, the pump body assembly being installed in the housing, and a crankshaft of the pump body assembly being connected to the motor.
It is a further object of the embodiments of the present application to provide a refrigerating and heating apparatus including a compressor as in the above embodiments.
The pump body subassembly that this application embodiment provided's beneficial effect lies in: compared with the prior art, the pump body assembly provided by the embodiment of the application has the advantages that the annular groove is formed in at least one of the main bearing and the auxiliary bearing, so that abrasion is reduced, and noise is reduced; and with the eccentric setting of the centre of rotation of at least one ring groove relative bent axle, can increase the intensity of the inside wall part area in ring groove to reduce bent axle to this part regional skew, and then reduce the skew of piston to this part region, so that the clearance is less between piston and the compression chamber inner surface, and then reduces high-pressure gas to the leakage of low-pressure gas, promotes compression efficiency.
The beneficial effect of the compressor that this application embodiment provided lies in: compared with the prior art, the compressor of this application embodiment has used the pump body subassembly of above-mentioned embodiment, and the high-pressure gas leaks to low-pressure gas in avoiding or reducing the compression chamber that can be better, and compression efficiency is higher.
The beneficial effect of the refrigeration equipment that this application embodiment provided lies in: compared with the prior art, the refrigerating and heating equipment of the embodiment of the application uses the compressor of the embodiment, has the technical effect of the compressor, and is not repeated herein.
Drawings
In order to more clearly illustrate the technical solutions in the embodiments of the present application, the drawings needed to be used in the embodiments or exemplary technical descriptions will be briefly described below, and it is obvious that the drawings in the following description are only some embodiments of the present application, and it is obvious for those skilled in the art to obtain other drawings without creative efforts.
Fig. 1 is a schematic cross-sectional view of a compressor according to an embodiment of the present disclosure;
FIG. 2 is a schematic cross-sectional view of the pump block assembly of FIG. 1;
FIG. 3 is a schematic cross-sectional view taken along line X-X in FIG. 2;
FIG. 4 is a schematic view of the ring groove portion of FIG. 3;
FIG. 5 is a schematic cross-sectional view taken along line Y-Y of FIG. 2;
FIG. 6 is a schematic cross-sectional view of the pump block assembly of FIG. 5 taken from the rotation angle of the piston to the eccentricity angle of the annular groove;
FIG. 7 is a graph illustrating the effect of the rotational angle of the piston on the torque of the crankshaft in the compressor according to an embodiment of the present invention;
fig. 8 is a schematic cross-sectional structure view of a pump body assembly in a compressor according to a second embodiment of the present application.
Wherein, in the drawings, the reference numerals are mainly as follows:
1A-a compressor; 1B-a compressor;
2-a housing; 3-an exhaust pipe; 4-a motor;
5A-a pump body assembly; 5B-a pump body assembly;
6-air suction pipe; 6A-an air suction pipe; 6B-suction pipe; 10-a spindle disc; 10A-main bearing; 10 a-main slide hole; 10C-vent hole; 10 c-centre of gyration; 11-a silencer; 12-a circular groove; 12A-ring bearing; 12 d-central axis; 14-a crankshaft; 15-a crankshaft; 15A-main shaft; 15B-eccentric shaft; 15C-auxiliary shaft; 16-a piston; 18-a slip sheet; 20-a cylinder; 20A-a compression chamber; 20 a-low pressure region; 20B-suction hole; 20 b-high pressure zone; 21-cylinder; 21A-a compression chamber; 22-a cylinder; 22A-a compression chamber; 25-a separator; 26A-a piston; 26B-a piston; 27A-eccentric shaft; 27B-eccentric shaft; 30-a countershaft disc; 30A-secondary bearing; 30 a-secondary slide hole.
Detailed Description
In order to make the technical problems, technical solutions and advantageous effects to be solved by the present application clearer, the present application is further described in detail below with reference to the accompanying drawings and embodiments. It should be understood that the specific embodiments described herein are merely illustrative of the present application and are not intended to limit the present application.
It will be understood that when an element is referred to as being "secured to" or "disposed on" another element, it can be directly on the other element or be indirectly on the other element. When an element is referred to as being "connected to" another element, it can be directly connected to the other element or be indirectly connected to the other element.
In the description of the present application, "a plurality" means two or more unless specifically limited otherwise. The meaning of "a number" is one or more unless specifically limited otherwise. The terms "first", "second" and "first" are used for descriptive purposes only and are not to be construed as indicating or implying relative importance or implicitly indicating the number of technical features indicated. Thus, a feature defined as "first" or "second" may explicitly or implicitly include one or more of that feature. The terms "center," "length," "width," "thickness," "upper," "lower," "front," "rear," "left," "right," "vertical," "horizontal," "top," "bottom," "inner," "outer," and the like are used in an orientation or positional relationship indicated in the drawings for convenience in describing the application and for simplicity in description, and are not intended to indicate or imply that the referenced device or element must have a particular orientation, be constructed and operated in a particular orientation, and thus are not to be considered limiting of the application.
In the description of the present application, it is to be noted that, unless otherwise explicitly specified or limited, the terms "mounted," "connected," and "connected" are to be construed broadly, e.g., as meaning either a fixed connection, a removable connection, or an integral connection; can be mechanically or electrically connected; either directly or indirectly through intervening media, either internally or in any other relationship. The specific meaning of the above terms in the present application can be understood by those of ordinary skill in the art as appropriate.
Reference throughout this specification to "one embodiment," "some embodiments," or "an embodiment" means that a particular feature, structure, or characteristic described in connection with the embodiment is included in one or more embodiments of the present application. Thus, appearances of the phrases "in one embodiment," "in some embodiments," "in other embodiments," or the like, in various places throughout this specification are not necessarily all referring to the same embodiment, but rather "one or more but not all embodiments" unless specifically stated otherwise. Furthermore, the particular features, structures, or characteristics may be combined in any suitable manner in one or more embodiments.
Rotary compressors generally comprise a casing and a pump block assembly mounted in the casing and supported by the casing. The pump body assembly generally comprises a cylinder, a piston, a crankshaft, a sliding vane, a main shaft disc and a secondary shaft disc; the piston is arranged in a compression cavity of the cylinder and is arranged on an eccentric shaft of the crankshaft, and the eccentric shaft drives the piston to revolve in the cylinder. A sliding sheet groove is formed in the air cylinder, the sliding sheet is slidably mounted in the sliding sheet groove and abuts against the piston, and the sliding sheet is pushed to reciprocate in the compression cavity when the piston revolves in the compression cavity. The main shaft disc and the auxiliary shaft disc cover two ends of the cylinder respectively. The main bearing is formed in the middle of the main shaft disc, and a main sliding hole is formed in the main bearing to support a main shaft of the crankshaft. An auxiliary bearing is formed in the middle of the auxiliary shaft disc, and an auxiliary sliding hole is formed in the auxiliary bearing to support an auxiliary shaft of the crankshaft. The bent axle often can have certain shake when rotating, sets up the ring groove like this on main bearing and auxiliary bearing, can reduce the friction with the bent axle, and then noise reduction and wearing and tearing.
When the pump body assembly works, the eccentric shaft of the crankshaft drives the piston to revolve in the compression cavity of the cylinder so as to compress gas. In order to ensure that the piston can rotate well in the compression cavity and avoid abrasion between the piston and the inner surface of the compression cavity, a certain gap is arranged between the outer peripheral surface of the piston and the inner surface of the compression cavity. And the piston is when the revolution, and gaseous pressure can crescent, leads to the piston to produce certain shake, and drives the bent axle shake, makes the inside wall in ring groove can take place to warp, and then leads to clearance between piston outer peripheral face and the compression intracavity surface to increase, and leads to high-pressure gas can leak low pressure gas, leads to the efficiency of compressor to descend, and compression efficiency reduces.
Based on this, this application provides a pump body subassembly of compressor to reduce the shake of piston revolution in the compression chamber, and then reduce the range that the clearance increases between the outer surface of piston and the compression intracavity surface that arouses by the shake, in order to reduce the leakage of high-pressure gas to low-pressure gas, promote the efficiency and the compression efficiency of compressor.
Please refer to fig. 1 to 7. Fig. 1 is a schematic cross-sectional structure diagram of the compressor of the present embodiment. The compressor is a single cylinder rotary compressor. Fig. 2 is a schematic cross-sectional structural view of the pump body assembly in the present embodiment. Fig. 3 is a schematic sectional view along line X-X in fig. 2. Fig. 4 is a schematic structural view of a circular groove portion in fig. 3. Fig. 5 is a schematic sectional view taken along the line Y-Y in fig. 2. FIG. 6 is a schematic cross-sectional view of the pump block assembly of FIG. 5 taken from the rotation angle of the piston to the eccentricity angle of the annular groove. Fig. 7 is a graph showing the effect of the rotational angle of the piston on the crank torque in the compressor of the present embodiment.
Referring to fig. 1 and 2, a compressor 1A provided in the present application will now be described. The compressor 1A comprises a machine shell 2, a motor 4 and a pump body assembly 5A, wherein the motor 4 and the pump body assembly 5A are installed in the machine shell 2, and the pump body assembly 5A is connected with the motor 4 so as to drive the pump body assembly 5A to operate through the motor 4.
The pump block assembly 5A includes a cylinder 20, a crankshaft 15, a piston 16, a slide 18, a main bearing 10A, and a secondary bearing 30A. Wherein: a compression chamber 20A is provided in the cylinder 20. The crankshaft 15 includes a main shaft 15A, an eccentric shaft 15B, and a sub shaft 15C connected in this order. In the compressor 1A, the main shaft 15A is connected to the motor 4 to drive the crankshaft 15 to rotate by the motor 4.
The piston 16 is disposed in the compression chamber 20A, and the piston 16 is mounted on an eccentric shaft 15B of the crankshaft 15 to rotate the piston 16 in the compression chamber 20A by the eccentric shaft 15B, so as to compress gas.
The main bearing 10A and the sub bearing 30A are respectively installed at both ends of the cylinder 20. The main bearing 10A is provided with a main slide hole 10A, and in use, a main shaft 15A of the crankshaft 15 is inserted into the main slide hole 10A, and the main shaft 15A of the crankshaft 15 is positioned by the main bearing 10A so that the main shaft 15A of the crankshaft 15 can rotate in the main slide hole 10A. The auxiliary bearing 30A is provided with an auxiliary slide hole 30A, and in use, the auxiliary shaft 15C of the crankshaft 15 is inserted into the auxiliary slide hole 30A, and the auxiliary shaft 15C of the crankshaft 15 is positioned by the auxiliary bearing 30A so that the auxiliary shaft 15C of the crankshaft 15 can rotate in the auxiliary slide hole 30A. Thus, the main bearing 10A and the sub bearing 30A support the main shaft 15A and the sub shaft 15C at both ends of the eccentric shaft 15B, respectively, and the main bearing 10A and the sub bearing 30A are installed at both ends of the cylinder 20, respectively, to position the crankshaft 15 and the cylinder 20 so that the eccentric shaft 15B rotates the piston 16 in the compression chamber 20A to compress the gas when the crankshaft 15 rotates.
One or both of the main bearing 10A and the auxiliary bearing 30A is/are provided with a circular groove 12 at one end close to the cylinder 20, that is, one end of the main bearing 10A close to the cylinder 20 is provided with a circular groove 12; or one end of the auxiliary bearing 30A close to the cylinder 20 is provided with a circular groove 12; or the main bearing 10A is provided with a circular groove 12 at one end close to the cylinder 20, and the sub bearing 30A is provided with a circular groove 12 at one end close to the cylinder 20.
In this embodiment, the annular groove 12 is provided at one end of the main bearing 10A close to the cylinder 20, and the annular groove 12 is provided at one end of the sub bearing 30A close to the cylinder 20, so that the wear between the main bearing 10A and the sub bearing 30A and the crankshaft 15, that is, the wear between the main bearing 10A and the main shaft 15A of the crankshaft 15, the wear between the sub bearing 30A and the sub shaft 15C of the crankshaft 15, and noise are reduced.
Referring to fig. 3 and 4, the central axis 12d of the at least one circular groove 12 is eccentrically disposed with respect to the rotation center 10c of the crankshaft 15. That is, when the main bearing 10A is provided with the annular groove 12 at an end close to the cylinder 20 and the sub bearing 30A is provided with the annular groove 12 at an end close to the cylinder 20, the central axis 12d of at least one annular groove 12 of the two annular grooves 12 is eccentrically disposed with respect to the revolution center 10c of the crankshaft 15. When the annular groove 12 is provided only at one end of the main bearing 10A close to the cylinder 20, the central axis 12d of the annular groove 12 is eccentric with respect to the rotation center 10c of the crankshaft 15. When the annular groove 12 is provided only at one end of the sub-bearing 30A close to the cylinder 20, the central axis 12d of the annular groove 12 is eccentrically provided with respect to the rotation center 10c of the crankshaft 15.
The inner side wall of the circular groove 12 is the side wall of the circular groove 12 close to the rotation center 10c of the crankshaft 15; for example, in the main bearing 10A, the inner wall of the annular groove 12 in the main bearing 10A means a side wall of the annular groove 12 on the side closer to the main shaft 15A of the crankshaft 15, and means a side wall of the annular groove 12 on the side closer to the main slide hole 10A. The inner side wall of the circular groove 12 may form a circular bearing 12A, that is, the circular bearing 12A is formed between the inner wall surface of the circular groove 12 on the main bearing 10A and the inner wall surface of the main sliding hole 10A; in the sub bearing 30A, a ring bearing 12A is formed between the inner wall surface of the ring groove 12 and the inner wall surface of the sub slide hole 30A.
By disposing the central axis 12d of at least one circular groove 12 eccentrically with respect to the rotation center 10c of the crankshaft 15, the thickness of the side of the circular bearing 12A formed by the inner side wall of the circular groove 12 is relatively large, and the rigidity and strength of the side of the circular bearing 12A are relatively large, so that the runout and the offset of the crankshaft 15 and the piston 16 to the side of the circular bearing 12A can be reduced when the crankshaft 15 rotates, and further, the gap between the piston 16 and the inner surface of the compression chamber 20A can be made small, so as to reduce the leakage of high-pressure gas to low-pressure gas.
Compared with the prior art, in the pump body assembly 5A provided by the embodiment of the application, the annular groove 12 is formed in at least one of the main bearing 10A and the auxiliary bearing 30A, so that abrasion is reduced, and noise is reduced; and at least one circular groove 12 is eccentrically arranged relative to the rotation center 10c of the crankshaft 15, so that the strength of the inner side wall partial area of the circular groove 12 can be increased, the deviation of the crankshaft 15 to the partial area is reduced, the deviation of the piston 16 to the partial area is further reduced, the gap between the piston 16 and the inner surface of the compression cavity 20A is smaller, the leakage of high-pressure gas to low-pressure gas is further reduced, and the compression efficiency is improved.
The compressor 1A provided by the embodiment of the application has the beneficial effects that: compared with the prior art, the compressor 1A of the embodiment of the application uses the pump body assembly 5A of the embodiment, is convenient to assemble, can be automatically assembled, and is high in assembling efficiency.
Referring to fig. 1 to 3, the pump body assembly 5A further includes a muffler 11, and the muffler 11 is mounted on the main bearing 10A to perform an exhaust silencing function. The cylinder 20 is provided with a suction hole 20B communicating with the compression chamber 20A, and the suction hole 20B is connected to the suction pipe 6. When the pump body assembly 5A works, the crankshaft 15 drives the piston 16 to revolve in the compression cavity 20A, the compression cavity 20A sucks gas from the air suction pipe 6, the piston 16 compresses the gas into high-pressure gas when rotating, the high-pressure gas is discharged to the silencer 11 from the exhaust hole 10C, the high-pressure gas discharged by the silencer 11 passes through the motor 4, and then the high-pressure gas is discharged out of the compressor 1A from the exhaust pipe 3. It is understood that the muffler 11 is not provided in the pump block assembly 5A, and the pump block assembly 5A can normally compress the gas.
Referring to fig. 1 to 4, the pump body assembly 5A includes a main shaft disk 10 and a sub shaft disk 30, wherein the main shaft disk 10 forms the main bearing 10A at a middle portion thereof. The auxiliary bearing 30A is formed in the middle of the auxiliary shaft disk 30. The main and sub-axial disks 10 and 30 cover opposite sides of the cylinder 20, respectively, to seal the compression chamber 20A of the cylinder 20.
Referring to fig. 2 to 4, the outer diameter of the circular ring bearing 12A is D, and the central axis of the circular ring bearing 12A is actually the central axis 12D of the corresponding circular ring groove 12. The offset distance of the central axis of the circular ring bearing 12A from the rotation center 10c of the crankshaft 15 is e, that is, the distance from the central axis 12d of the circular ring groove 12 to the rotation center 10c of the crankshaft 15 is e, that is, the eccentricity of the circular ring groove 12 is e.
In one embodiment, when the main bearing 10A and the sub bearing 30A are each provided with the annular groove 12, and both the annular grooves 12 are eccentrically provided with respect to the rotation center 10c of the crankshaft 15, the eccentric amounts e of both the annular grooves 12 are equal, that is, the eccentric distances of both the annular grooves 12 with respect to the rotation center 10c of the crankshaft 15 are equal. Therefore, the main shaft 15A and the auxiliary shaft 15C of the crankshaft 15 can be synchronously positioned by the circular ring bearings 12A corresponding to the two circular ring grooves 12, so that the jumping and the deviation of the crankshaft 15 are better avoided, and further, the clearance between the piston 16 and the inner surface of the compression cavity 20A can be better smaller, so that the leakage of high-pressure gas to low-pressure gas is better reduced.
Referring to fig. 3 to 6, the eccentric angle θ 1 of the circular groove 12 is: the direction 12E from the rotation center 10c of the crankshaft 15 to the central axis 12d of the circular groove 12 is an angle along the revolving direction of the piston 16 with the center line of the reciprocating motion of the slide 18 as a reference. That is, the eccentric angle θ 1 of the circular groove 12 is a deflection angle of the central axis 12d of the circular groove 12 with respect to the rotation center 10c of the crankshaft 15 and in the revolving direction of the piston 16 with respect to the center line of the reciprocating motion of the vane 18.
The reciprocating central line of the sliding piece 18 is used as a reference, the revolving angle of the piston 16 in the compression cavity 20A is theta, and when the piston 16 rotates to a position where the sliding piece 18 is attached to the compression cavity 20A, namely the sliding piece 18 is located at the top dead center, and the length of the sliding piece 18 extending into the compression cavity 20A is shortest, the revolving angle theta of the piston 16 is 0 deg.
In one embodiment, referring to fig. 3 to 6, the eccentric angle θ 1 of the circular groove 12 ranges from 0 to 90 degrees. Setting the eccentric angle θ 1 of the circular groove 12 to 0 ° -90 °, that is, the circular groove 12 is eccentrically disposed toward the side of the center line of the reciprocating motion of the sliding vane 18 close to the suction hole 20B, so that the strength of the circular bearing 12A located on the side of the center line of the reciprocating motion of the sliding vane 18 close to the suction hole 20B is relatively large, and thus when the piston 16 rotates in the compression chamber 20A and the piston 16 rotates to the side of the center line of the reciprocating motion of the sliding vane 18 close to the exhaust hole 10C to compress the gas into high-pressure gas, it is better to prevent the piston 16 from shifting toward the side of the center line of the reciprocating motion of the sliding vane 18 close to the suction hole 20B, so that when the piston 16 rotates in the compression chamber 20A at the side of the center line of the reciprocating motion of the sliding vane 18 close to the exhaust hole 10C, the gap between the piston 16 and the inner surface of the compression chamber 20A is kept small, so as to better reduce or prevent the high-pressure gas from leaking to low-pressure gas, to improve compression efficiency.
In order to ensure that the piston 16 can rotate well in the compression chamber 20A and avoid abrasion between the piston 16 and the inner surface of the compression chamber 20A, a certain clearance is provided between the outer peripheral surface of the piston 16 and the inner surface of the compression chamber 20A. However, due to the clearance, the high-pressure gas in the compression chamber 20A leaks into the low-pressure gas when the compressor 1A is operated.
In one embodiment, referring to fig. 3 to 6, the center of the compression chamber 20A is eccentrically located toward one side of the rotational center 10c of the crankshaft 15. The center of the compression chamber 20A is aligned toward the rotation center 10c of the crankshaft 15. The distance by which the center of compression chamber 20A is offset from the center of rotation 10c of crankshaft 15 is referred to as the amount of misalignment. Since the center of the compression chamber 20A is shifted toward one side of the rotation center 10c of the crankshaft 15, when the eccentric shaft 15B of the crankshaft 15 drives the piston 16 to rotate in the compression chamber 20A for one revolution, the clearance between the piston 16 and the inner surface of the compression chamber 20A has a minimum clearance throughout the compression chamber 20A, and at this time, the minimum clearance position Cm is formed at the position where the distance from the inner surface of the compression chamber 20A to the outer peripheral surface of the piston 16 is minimum, that is, when the piston 16 revolves in the compression chamber 20A and the piston 16 reaches the minimum clearance position Cm, the distance of the minimum clearance between the piston 16 and the compression chamber 20A is smaller relative to the minimum clearance at the other position where the piston 16 rotates to the compression chamber 20A, and generally, when the piston 16 revolves to the opposite direction of the center shift of the compression chamber 20A, the clearance between the piston 16 and the inner surface of the compression chamber 20A is minimum, the minimum clearance position Cm is formed at the minimum distance from the outer circumferential surface of the piston 16 to the inner surface of the compression chamber 20A.
The direction from the rotation center 10c of the crankshaft 15 to the minimum clearance position Cm is defined as an angle θ 2 of the center of alignment of the minimum clearance position Cm with respect to the center line of the reciprocating movement of the slide piece 18 and the angle in the revolving direction of the piston 16. When the piston 16 is rotated to the center adjustment angle θ 2, that is, when the revolution angle θ of the piston 16 is equal to θ 2, the clearance between the inner surface of the compression chamber 20A and the outer peripheral surface of the piston 16 is minimum, that is, when the revolution angle θ of the piston 16 is equal to θ 2, the piston 16 reaches the minimum clearance position Cm. The center adjusting angle θ 2 of the minimum clearance position Cm also means: the rotation center 10c of the crankshaft 15 is oriented to the minimum clearance position Cm with respect to the center line of the reciprocating movement of the slide plate 18, and is deflected by an angle in the revolving direction of the piston 16. When the revolution angle θ of the piston 16 is θ 2, the clearance between the inner surface of the compression chamber 20A and the outer peripheral surface of the piston 16 is minimized, so that the leakage of high-pressure gas to low-pressure gas can be avoided or reduced, and the compression efficiency of the pump block assembly 5A can be improved.
Referring to fig. 5, when the revolution angle θ of the piston 16 is the center adjusting angle θ 2, the piston 16 reaches the minimum clearance position Cm, the inner surface of the compression chamber 20A and the reverse acting force of the high-pressure gas on the piston 16 are also maximized, and accordingly, the pressure difference (Pd-Ps) between the high-pressure gas compression Pd and the low-pressure gas Ps is maximized, that is, the pressure difference (Pd-Ps) between the high-pressure region 20b and the low-pressure region 20A in the compression chamber 20A is maximized, so that the piston 16 and the crankshaft 15 reversely bounce or shift to the minimum clearance position Cm along the radial direction of the compression chamber 20A, and accordingly, the annular bearing 12A deforms, so that the clearance between the piston 16 and the compression chamber 20A at the minimum clearance position Cm is increased, and the high-pressure gas leaks to the low-pressure gas.
In one embodiment, referring to fig. 3 to 6, the direction from the central axis 12d of the circular groove 12 to the rotation center 10c of the crankshaft 15 is from-50 degrees to 40 degrees relative to the direction from the minimum clearance position Cm to the rotation center 10c of the crankshaft 15 and along the revolving direction of the piston 16. That is, the included angle between the direction from the central axis 12d of the circular groove 12 to the rotation center 10c of the crankshaft 15 and the direction opposite to the minimum clearance position Cm along the radial direction of the compression cavity 20A is in the range of-50 ° to 40 °, so that the rigidity and strength of the circular bearing 12A on the side away from the minimum clearance position Cm are relatively large, when the piston 16 and the crankshaft 15 jump or shift to the side away from the minimum clearance position Cm, the circular bearing 12A can abut against the crankshaft 15, thereby reducing or avoiding the piston 16 and the crankshaft 15 from jumping or shifting, and further making the clearance between the piston 16 and the compression cavity 20A at the minimum clearance position Cm and the vicinity of the minimum clearance position Cm small, thereby avoiding or reducing the leakage of high-pressure gas to low-pressure gas.
In one embodiment, referring to fig. 3 to 6, the central axis 12d of the circular groove 12 is located on an extension line from the minimum clearance position Cm to the rotation center 10c of the crankshaft 15, that is, the central axis 12d of the circular groove 12 is located in a direction away from the minimum clearance position Cm along the radial direction of the compression cavity 20A, so that when the piston 16 rotates to the aligning angle θ 2, the reverse acting force of the inner surface of the compression cavity 20A and the high-pressure gas to the piston 16 is also maximum, and the rigidity and strength of the corresponding circular bearing 12A away from the minimum clearance position Cm are also maximum, which can better reduce or prevent the piston 16 and the crankshaft 15 from jumping or shifting, and further make the clearance between the piston 16 and the compression cavity 20A at the minimum clearance position Cm and near the minimum clearance position Cm smaller to prevent or reduce the leakage of the high-pressure gas to the low-pressure gas.
Referring to fig. 6, when the revolution angle θ of the piston 16 is the eccentric angle θ 1, the rotation angle of the piston 16 is small, the pump body assembly 5A is in the suction stroke of the low-pressure gas, and the suction and compression strokes of the compression chamber 20A are shifted from the low pressure to the high pressure. The piston 16 is at its greatest distance from the minimum clearance position Cm and the high side pressure of the compression chamber 20A is small. Therefore, the high-low pressure difference (Pd-Ps) is minimized, that is, the pressure difference (Pd-Ps) between the high-pressure region 20b and the low-pressure region 20A in the compression chamber 20A is minimized, the amount of gas leaked into the low-pressure chamber is small, and the re-expansion loss of the leaked gas is also small.
Referring to fig. 2 to 4, since the circular groove 12 is eccentrically disposed, the minimum thickness of the circular bearing 12A is W1, and the maximum thickness of the circular bearing 12A is W2. The thickness of the portion 12b of the inner surface of the circular ring bearing 12A farthest from the center axis 12d of the circular ring groove 12 in the eccentric direction of the circular ring groove 12 is the minimum thickness W1 of the circular ring bearing 12A; the thickness of the closest part 12c of the inner surface of the ring bearing 12A to the central axis 12d of the ring groove 12 is the maximum thickness W2 of the ring bearing 12A, W1 is less than W2, the rigidity of the ring bearing 12A at the thickness W2 is maximum, the deformation is minimum correspondingly, the crankshaft 15 and the piston 16 can be prevented from deforming to the position where the thickness W2 of the ring bearing 12A is, and the leakage of high-pressure gas to low-pressure gas is avoided.
In one embodiment, the rotation center 10c of the crankshaft 15 is to the minimum clearance position Cm with the center line of reciprocation of the vane 18 as a reference, and the angle in the revolving direction of the piston 16 ranges from 180 degrees to 270 degrees; that is, the range of the center adjusting angle θ 2 of the minimum clearance position Cm is 180 degrees to 270 degrees, so that when the piston 16 rotates in the compression chamber 20A on the side of the reciprocating center line of the sliding vane 18 close to the exhaust hole 10C, the clearance between the piston 16 and the inner surface of the compression chamber 20A is kept small, thereby better reducing or avoiding the leakage of high-pressure gas to low-pressure gas, and improving the compression efficiency.
In one embodiment, based on the center line of the reciprocating motion of the sliding vane 18, the rotation center 10C of the crankshaft 15 is located at the minimum clearance position Cm, and the included angle along the revolving direction of the piston 16 is 230 degrees, that is, the aligning angle θ 2 of the minimum clearance position Cm is 230 °, at which time the high-pressure gas is ready to enter the exhaust hole 10C for exhaust, the pressure of the corresponding high-pressure gas is also maximum, and the aligning angle θ 2 is 230 °, so that the piston 16 revolves to the aligning angle θ 2, and the minimum clearance between the piston 16 and the inner surface of the compression cavity 20A is minimized, thereby better reducing the leakage of the high-pressure gas to the low-pressure gas. It can be understood that, according to the different angles when the piston 16 revolves to exhaust in different cylinders 20, the corresponding aligning angle θ 2 can also be set differently, generally, the aligning angle θ 2 is set to the angle when the piston 16 revolves just to exhaust, and the minimum clearance between the piston 16 and the inner surface of the compression cavity 20A is minimized, so as to better reduce the leakage of the high-pressure gas to the low-pressure gas.
In the above embodiment, when the center axis 12d of the annular groove 12 is located in the direction away from the minimum clearance position Cm in the radial direction of the compression chamber 20A, that is, when the eccentric angle θ 1 of the annular groove 12 is 50 °, when the piston 16 rotates to the center angle θ 2, the reverse acting force of the inner surface of the compression chamber 20A and the high-pressure gas on the piston 16 is also the largest, the rigidity and strength of the corresponding annular bearing 12A away from the minimum clearance position Cm are also the largest, the run-out or the deflection of the piston 16 and the crankshaft 15 can be better reduced or avoided, and further, the clearance between the piston 16 and the compression chamber 20A in the vicinity of the minimum clearance position Cm and the minimum clearance position Cm is smaller, so as to avoid or reduce the leakage of the high-pressure gas to the low-pressure gas.
In one embodiment, referring to fig. 2 and 3, the end surface of the annular bearing 12A is also an end surface of the inner side wall of the annular groove 12 close to the cylinder 20. The end surface 12A of the outer side wall of the circular groove 12 is spaced from the end surface of the circular bearing 12A, and the end surface of the circular bearing 12A is located on the side of the end surface 12A of the outer side wall of the circular groove 12 away from the cylinder 20, that is, along the axial direction of the compression chamber 20A: the distance from the end face of the circular ring bearing 12A to the cylinder 20 is greater than the distance from the end face 12A of the outer side wall of the circular ring groove 12 to the cylinder 20, namely, the axial length of the outer wall surface of the circular ring groove 12 is greater than the axial length of the inner wall surface of the circular ring groove 12, so that the cross section of the circular ring groove 12 along the central shaft 12d is L-shaped, the abrasion between the circular ring bearing 12A and the crankshaft 15 can be better reduced, and the noise is reduced. For the main bearing 10A, the outer side wall of the circular groove 12 herein actually refers to a portion of the main shaft disk 10 between the outer wall surface of the circular groove 12 and the outer peripheral surface of the main shaft disk 10; in the auxiliary bearing 30A, the outer wall of the annular groove 12 herein means a portion of the auxiliary disk 30 between the outer wall surface of the annular groove 12 and the outer peripheral surface of the auxiliary disk 30.
Referring to fig. 2, 3 and 7, fig. 7 shows the compressor 1A of the present embodiment applied to an air conditioner, and employs a refrigerant R32(R32, chemical name is difluoromethane, molecular formula is CH2F 2). The abscissa of the figure is the revolution angle θ of the piston 16 in the compression chamber 20A, the ordinate is the rotational moment Tr of the crankshaft 15, and the graph is the change in the rotational moment Tr of the crankshaft 15 when the piston 16 revolves in the compression chamber 20A. Compression in the figure and Discharge in the figure. When the revolution angle theta of the piston 16 is 0 DEG, the refrigerant sucked into the compression cavity 20A is compressed by the revolution of the piston 16; when θ is 90 °, the rotational torque Tr of the crankshaft 15 is about 0.1; the rotation torque Tr of the crankshaft 15 is 0.5 when θ is 180 °; when θ is 230 °, Tr is 1.0, and the rotational torque Tr of the crankshaft 15 is maximum. When θ is 230 °, high-pressure gas is discharged from the gas discharge hole 10C into the casing, so that the high-pressure gas pressure can be maintained, but the rotation torque Tr rapidly decreases, and Tr is about 0 when θ of the vane 18 reaches the top dead center is 360 °. The compressor 1A rotates at 20 to 100rps, and therefore the rotational torque change is also synchronized with the rotational speed.
Referring to fig. 3 and 4, fig. 3 is a schematic structural diagram of the piston 16 rotating to the aligning angle θ 2. At this time, the rotational torque Tr of the crankshaft 15 is at the maximum state where θ becomes 230 °, and the piston 16 is at the minimum clearance position. Of course, the pressure change and the rotation torque curve of the crankshaft 15 may be changed in some ways depending on the design of the air conditioner or the refrigeration equipment in which the compressor 1A is installed, the type of the refrigerant used, and the like.
From the above-described changes in the operating conditions, it can be seen that the center adjusting angle θ 2 of the minimum clearance position Cm between the outer peripheral surface of the piston 16 and the inner surface of the compression chamber 20A varies from 180 ° to 270 °, and therefore the optimum eccentric angle θ 1 of the annular groove 12 can range from 0 ° to 90 °.
Referring to fig. 8, fig. 8 is a schematic cross-sectional view of a pump body assembly of the compressor according to the present embodiment. The structure of the present embodiment is a modification of the structure of fig. 1, in the present embodiment, the pump body assembly 5B is a double-cylinder structure, so as to form a compressor 1B double-cylinder compressor, specifically, the pump body assembly 5B includes a cylinder 21 and a cylinder 22, a partition 25 is disposed between the cylinder 21 and the cylinder 22, a compression cavity 21A is disposed in the cylinder 21, a compression cavity 22A is disposed in the cylinder 22, the crankshaft 14 has an eccentric shaft 27A and an eccentric shaft 27B, a piston 26A is mounted on the eccentric shaft 27A to drive the piston 26A to revolve in the compression cavity 21A through the eccentric shaft 27A, and a piston 26B is mounted on the eccentric shaft 27B to drive the piston 26B to revolve in the compression cavity 22A through the eccentric shaft 27B. A suction pipe 6A connected to the cylinder 21 and a suction pipe 6B connected to the cylinder 22 are installed on the casing 2 of the compressor 1B to allow the cylinder 21 and the cylinder 22 to suck air.
The other structure of the pump body assembly 5B of the present embodiment is the same as the structure of the pump body assembly 5A shown in fig. 1, and is not described herein again.
It will be understood that the pump block assembly may also comprise a greater number of cylinders in each of which a piston is mounted, the crankshaft comprising an eccentric shaft corresponding to the piston, with a partition being provided between two adjacent cylinders to form a multi-cylinder pump block, the corresponding compressor forming a multi-cylinder compressor.
The pump body assembly of the embodiment of the application can improve the re-expansion loss caused by the leakage of the high-pressure gas in the compression cavity, and further can improve the compression efficiency and the energy efficiency of the compressor. In addition, the circular groove is eccentrically arranged, so that the design and the manufacture are convenient, and the manufacturing cost cannot be increased.
Embodiments of the present application further provide a refrigeration and heating apparatus, including a compressor according to any one of the above embodiments. The refrigeration and heating equipment uses the compressor of the embodiment, has the technical effects of the compressor of the embodiment, and is not described again here.
The cooling and heating device in the embodiment of the application may be a device only for cooling, a device only for heating, or a device both for cooling and heating.
The present invention is not intended to be limited to the particular embodiments shown and described, but is to be accorded the widest scope consistent with the principles and novel features herein disclosed.
Claims (14)
1. A pump body assembly, comprising:
the air cylinder is internally provided with a compression cavity;
the crankshaft comprises a main shaft, an eccentric shaft and an auxiliary shaft which are connected in sequence;
the piston is arranged on the eccentric shaft and revolves in the compression cavity;
the sliding sheet is abutted against the piston and reciprocates in the compression cavity;
the main bearing is arranged at one end of the air cylinder, and a main sliding hole which is matched and sleeved on the main shaft is arranged in the main bearing; and the number of the first and second groups,
the auxiliary bearing is arranged at the other end of the cylinder, and an auxiliary sliding hole which is matched and sleeved on the auxiliary shaft is formed in the auxiliary bearing;
one end of one or two of the main bearing and the auxiliary bearing, which is close to the air cylinder, is provided with a circular groove, and the central shaft of at least one circular groove is eccentrically arranged relative to the rotation center of the crankshaft.
2. The pump body assembly of claim 1, wherein: and taking the reciprocating central line of the sliding sheet as a reference, wherein the central axis of the annular groove is relative to the rotation center of the crankshaft, and the range of an eccentric angle theta 1 along the revolution direction of the piston is 0-90 degrees.
3. The pump body assembly of claim 2, wherein: the center of the compression chamber is eccentrically disposed with respect to the rotation center of the crankshaft, and a minimum clearance position is formed at a position where a clearance between the inner surface of the compression chamber and the outer circumferential surface of the piston is at a minimum.
4. The pump body assembly of claim 3, wherein: the direction from the central axis of the circular groove to the rotation center of the crankshaft is opposite to the direction from the minimum clearance position to the rotation center of the crankshaft, and the included angle range along the revolution direction of the piston is-50 degrees to 40 degrees.
5. The pump body assembly of claim 4, wherein: and the central shaft of the annular groove is positioned on an extension line from the position of the minimum clearance to the rotation center of the crankshaft.
6. The pump body assembly of claim 3, wherein: and taking the reciprocating central line of the sliding sheet as a reference, and setting the range of an included angle from the rotation center of the crankshaft to the position of the minimum clearance and along the revolution direction of the piston to be 180-270 degrees.
7. The pump body assembly of claim 6, wherein: and taking the reciprocating central line of the sliding sheet as a reference, and enabling the rotation center of the crankshaft to reach the minimum clearance position, wherein the included angle along the revolution direction of the piston is 230 degrees.
8. The pump body assembly of any one of claims 1-7, wherein: the main bearing and the auxiliary bearing are both provided with the annular grooves.
9. The pump body assembly of claim 8, wherein: the central shafts of the circular grooves on the main bearing and the auxiliary bearing are eccentrically arranged relative to the revolution center of the crankshaft.
10. The pump body assembly of claim 9, wherein: the eccentric distances of the two circular ring grooves relative to the rotation center of the crankshaft are equal.
11. The pump body assembly of any one of claims 1-7, wherein: the inner side wall of the circular groove forms a circular bearing; in the axial direction of the compression chamber: the distance from the end face of the circular ring bearing to the cylinder is larger than the distance from the end face of the outer side wall of the circular ring groove to the cylinder.
12. The pump body assembly of any one of claims 1-7, wherein: the eccentric shaft is arranged in the cylinder, the eccentric shaft is arranged on the eccentric shaft, and the eccentric shaft is arranged on the eccentric shaft.
13. A compressor, includes the casing and installs in the motor in the casing, its characterized in that: further comprising a pump block assembly according to any one of claims 1 to 12 mounted in the housing, the crankshaft of the pump block assembly being connected to the motor.
14. A refrigerating and heating apparatus characterized by: comprising a compressor as claimed in claim 13.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CN202111620461.5A CN114320915A (en) | 2021-12-27 | 2021-12-27 | Pump body subassembly, compressor and refrigeration equipment of heating |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CN202111620461.5A CN114320915A (en) | 2021-12-27 | 2021-12-27 | Pump body subassembly, compressor and refrigeration equipment of heating |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CN114320915A true CN114320915A (en) | 2022-04-12 |
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Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CN202111620461.5A Pending CN114320915A (en) | 2021-12-27 | 2021-12-27 | Pump body subassembly, compressor and refrigeration equipment of heating |
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| Country | Link |
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| CN (1) | CN114320915A (en) |
Cited By (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CN114753989A (en) * | 2022-05-16 | 2022-07-15 | 珠海凌达压缩机有限公司 | Piston type compressor and refrigeration equipment applying same |
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| JPH10141271A (en) * | 1996-11-01 | 1998-05-26 | Daikin Ind Ltd | Rotary compressor |
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| CN210050033U (en) * | 2019-06-28 | 2020-02-11 | 广东美芝制冷设备有限公司 | Rotary compressor and heat exchange equipment |
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| JPH0169194U (en) * | 1987-09-14 | 1989-05-08 | ||
| JPH04143483A (en) * | 1990-10-05 | 1992-05-18 | Daikin Ind Ltd | Compressor with rolling piston |
| JPH10141271A (en) * | 1996-11-01 | 1998-05-26 | Daikin Ind Ltd | Rotary compressor |
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