JP4674466B2 - Compressor - Google Patents

Description

  The present invention relates to a main shaft and a main bearing that can improve reliability in a compressor used in a domestic refrigerator-freezer.

  In recent years, for home refrigerators, in addition to improving basic performance such as freshness, energy saving, and quietness, it has been required to improve usability and storage. The trend is to increase. Under such circumstances, downsizing of the compressor, which is one of the devices governing the size of the machine room of a home refrigerator, is strongly desired.

  As a conventional compressor, there is one in which an electric element composed of a stator and a rotor and a compression element driven by the electric element are housed in an airtight container (for example, see Patent Document 1).

  The conventional compressor will be described below with reference to the drawings. In the following description, the upper and lower relationships are based on a state where the compressor is installed in a normal posture.

  FIG. 13 is a longitudinal sectional view of a conventional compressor, and FIG. 14 is a characteristic diagram showing the behavior of a conventional main shaft.

  The sealed container 1 is filled with the refrigerant 21 and stores the oil 2.

  The electric element 5 includes a stator 3 fixed below the cylinder block 10 and connected to an inverter drive circuit (not shown), and a rotor 4 containing a permanent magnet and fixed below the main shaft 7. .

  The compression element 6 will be described below.

  The crankshaft 9 extends in the vertical direction and includes a main shaft 7 and an eccentric shaft 8. The cylinder block 10 has a bore 13 that forms a compression chamber 12 together with a main bearing 11 that rotatably supports the main shaft 7.

  The piston 14 is inserted into the bore hole 13 so as to be reciprocally movable. The piston pin 15 has a substantially cylindrical shape, is disposed in parallel with the eccentric shaft 8, and is locked to a piston pin hole 16 formed in the piston 14 so as not to rotate. The connecting means 17 connects the eccentric shaft 8 and the piston 14 via the piston pin 15.

  The main shaft 7 includes a spiral groove 18 formed on a sliding surface with the main bearing 11, and an oil pump hole 19 that is connected to the lower end of the spiral groove 18 through a communication hole 20 and opens to the oil 2. Yes.

  The crankshaft 9 and the cylinder block 10 are both made of gray cast iron (FC200), which is a cast iron material that is inexpensive and easy to process.

  About the compressor comprised as mentioned above, the operation | movement is demonstrated below.

  When the electric element 5 is energized, the rotor 4 rotates and the crankshaft 9 rotates accordingly. The main shaft 7 rotates in the main bearing 11, while the rotational movement of the eccentric shaft 8 is transmitted to the piston 14 through the connecting means 17 and the piston pin 15, and the piston 14 reciprocates in the bore hole 13. Due to the reciprocating motion of the piston 14, the refrigerant 21 in the sealed container 1 is sucked into the compression chamber 12 and then compressed and discharged to the outside of the sealed container 1, and so-called suction stroke and compression stroke are repeated.

The oil 2 is sucked into the oil pump hole 19 by the centrifugal force accompanying the rotation of the crankshaft 9, guided upward from the spiral groove 18 through the communication hole 20, and injected from the upper end of the eccentric shaft 8. The oil 2 lubricates sliding portions such as the connecting means 17 and the piston pin 15 and the piston 14 and the bore hole 13.
JP 2000-145637 A

  However, in the above-described conventional configuration, the sliding portion between the main shaft 7 and the main bearing 11 may be worn due to one piece. During the compression stroke, a compressive load P is applied to the eccentric shaft 8 via the connecting means 17, and the main shaft 7 is inclined in the main bearing 11 and contacts in a local region above and below the main bearing 11. Sliding in a so-called one-sided state.

  According to the inventors' actual machine test, especially when the piston 14 is located near the top dead center and the maximum value of the compression load P is applied, the upper side (point U in FIG. 14) and the lower side (point U in FIG. 14). The wear at the main shaft 7 portion that contacts the point M in FIG. 14 is significant compared to other portions, and is lower than the main shaft 7 portion that contacts the upper portion of the main bearing 11 (point U in FIG. 14). It has been confirmed that the wear at the portion of the main shaft 7 in contact with (point M in FIG. 14) is greater.

  In addition, when the length of the main bearing 11 is further shortened in response to the recent demand for compressor downsizing, the contact between the main shaft 7 and the main bearing 11 due to the moment relationship even if the compression load P is the same as the conventional one. It has also been confirmed that the load increases, and in particular, wear at the portion of the main shaft 7 that contacts the lower portion of the main bearing 11 (point M in FIG. 14) becomes larger.

  When sliding in such a state where it comes into contact with each other, the contact surface is a very narrow region, and in addition, the load concentrates in the narrow region, so that the lubricating action by the oil 2 is not sufficiently exhibited, Abnormal wear may occur, or wear powder generated between the sliding surfaces of the main shaft 7 and the main bearing 11 may be caught and lock (unrotatable) may occur, which may hinder long-term reliability. There was a drawback of coming.

  Further, measures for preventing wear of the main shaft 7 and the main bearing 11 have been studied from various viewpoints such as a mechanism surface and a material surface, but this leads to a complicated configuration and high cost, or downsizing of the compressor. There were drawbacks such as being unable to cope with.

  SUMMARY OF THE INVENTION The present invention solves the above-described conventional problems, and an object of the present invention is to provide a highly reliable and inexpensive compressor that prevents wear due to contact between the main shaft 7 and the main bearing 11.

In order to solve the above problems, the compressor of the present invention, the rigidity ratio Gb / Gc and flexural rigidity Gb and flexural rigidity Gc of the principal axes of the main bearing in the thickest thin straight portion of the main bearing, 1 .8 or more and 4.6 or less , when the bending rigidity of the main shaft is remarkably lowered with respect to the main bearing, when the main shaft is tilted and comes into contact with the main bearing during the compression stroke Since the main shaft is curved starting from the portion that contacts the lower portion of the main bearing and conforms to the inner diameter of the main bearing, the contact area between the main shaft and the main bearing is increased, and wear caused by one piece is reduced.

  Since the compressor according to the present invention reduces wear due to the contact between the main shaft and the main bearing, it is possible to provide a highly reliable and inexpensive compressor.

According to the first aspect of the present invention, oil is stored in a sealed container, and an electric element composed of a stator and a rotor and a compression element driven by the electric element are accommodated, and the compression element extends in the vertical direction. A crankshaft having a main shaft and a cylinder block having a main bearing that supports the main shaft, and the bending rigidity Gb of the main bearing and the bending rigidity Gc of the main shaft in the thinnest straight portion of the main bearing. The rigidity ratio Gb / Gc is set to 1.8 or more and 4.6 or less. By significantly reducing the bending rigidity of the main shaft relative to the main bearing, the main shaft is inclined by receiving a compressive load during the compression stroke. When it comes into contact with the main bearing, the main shaft is curved starting from the part that contacts the lower part of the main bearing and conforms to the inner diameter of the main bearing, increasing the contact area between the main shaft and the main bearing and reducing wear due to contact with the main bearing. Reliable, cheap Inexpensive compressors can be provided.

The invention according to claim 2, in which Oite to the invention of claim 1, the ratio L / D of the main bearing length L and internal diameter D, and 2.5 or less 1.9 or more, the main axis of the piece To provide a highly reliable and inexpensive compressor that can reduce the size of the compressor by shortening the length of the main bearing and the main shaft while significantly reducing the wear reduction effect due to hitting. Can do.

Invention according to claim 3, Oite to the invention of claim 1 or claim 2, the material of the main bearing, a spherical graphite cast iron or gray cast iron, which was a material of the main shaft and the gray cast iron, current production It is not necessary to significantly change the line, and it is possible to provide a compressor that is excellent in productivity, high in reliability, and inexpensive.

The invention according to claim 4, Oite claims 1 to any one of the invention of claim 3, which has an upper end in the cylindrical hole of the oil pump hole extending bored easily just hole processing The bending rigidity of the main shaft can be reduced, and in addition, the diameter and length of the cylindrical hole can be changed according to the compression load conditions required for the compressor, so it is extremely versatile and highly reliable. A high and inexpensive compressor can be provided.

The invention according to claim 5, Oite claim 1 the invention of any one of claims 4, intended to be operated at operating frequencies including frequencies below at least utility frequency, the input of the compressor In addition, the wear of the main shaft and the main bearing can be prevented, and a highly reliable and inexpensive compressor can be provided.

(Embodiment 1)
FIG. 1 is a longitudinal sectional view of a compressor according to Embodiment 1 of the present invention, and FIG. 3 is a characteristic diagram showing the behavior of the main shaft in the same embodiment, FIG. 4 is a characteristic diagram showing a measurement result of the main shaft wear amount in the same embodiment, and FIG. 5 is a diagram showing the rigidity ratio Gb / Gc. FIG. 6 is a characteristic diagram showing the correlation result between the main bearing length / inner diameter ratio L / D and the main shaft wear amount.

  The hermetic container 101 is filled with isobutane (R600a) as the refrigerant 121 and stores a relatively low viscosity mineral oil as the oil 102.

  The electric element 105 includes a stator 103 fixed below the cylinder block 110 and connected to an inverter drive circuit (not shown), and a rotor 104 containing a permanent magnet and fixed below the main shaft 107. An electric motor for driving the inverter is formed, and is driven at a plurality of operation frequencies including an operation frequency (for example, 1500 r / min) lower than the commercial power supply frequency by the inverter drive circuit.

  The stator 103 is a salient pole concentrated winding. The salient pole concentrated winding type winding has an iron core having a portion (a salient pole) projecting radially toward the center of the rotor 104, and a covered wire is concentrated around the salient pole of the iron core. It is attached. Since the salient pole concentrated winding has very little protrusion from the iron core, the specification is such that the total height of the electric element 105 is lower than that of other winding types (for example, distributed winding type).

  The compression element 106 will be described below.

  The crankshaft 109 extends in the vertical direction and includes a main shaft 107 and an eccentric shaft 108. The cylinder block 110 has a bore hole 113 that forms a compression chamber 112 together with a main bearing 111 that rotatably supports the main shaft 107.

  The piston 114 is reciprocally inserted into the bore hole 113. The piston pin 115 has a substantially cylindrical shape, is disposed in parallel with the eccentric shaft 108, and is non-rotatably locked in a piston pin hole 116 formed in the piston 114. The connecting means 117 connects the eccentric shaft 108 and the piston 114 via the piston pin 115.

  The inner diameter D of the main bearing 111 is set to 18 mm which can effectively suck up the oil 102 by effectively applying centrifugal force even at low speed operation (for example, the operation frequency is 1500 r / min).

  The clearance between the outer diameter of the main shaft 107 and the inner diameter D of the main bearing 111 is within a range of 10 to 30 μm in consideration of machining accuracy, twisting due to a change in diameter during heating, or prevention of rattling during rotation of the main shaft 107. It is. The main bearing 111 has a length L of 40 mm.

  The main shaft 107 is connected to the spiral groove 118 formed on the sliding surface with the main bearing 111 and the lower end of the spiral groove 118 through the communication hole 120, and is inclined upward from the bottom opened to the oil 202. An oil pump hole 119 is provided.

The crankshaft 109 having the main shaft 107 is made of gray cast iron (FC200) having a Young's modulus of 73.5 to 127 × 10 ⁻⁹ N / m ² . The cylinder block 110 having one main bearing 111 is formed of spheroidal graphite cast iron (FCD500) having a Young's modulus of 140 to 170 × 10 ⁻⁹ N / m ² .

  Due to the structure of the compressor, the straight portion 123 of the main bearing 111 surrounded by the recess 122 formed on the upper end surface of the rotor 104 corresponds to the thinnest portion. The minimum thickness below the main bearing 111 is set to 2.5 mm in consideration of processing, assembly tolerances, impact cracking, brittle cracking due to metal fatigue during long-term operation, and the like. Therefore, the outer diameter of the straight portion 123 of the main bearing 111 is 23 mm.

  Here, the bending rigidity indicates the magnitude of deformation resistance to bending, and is represented by the product of Young's modulus and cross-sectional second moment.

  From the dissimilar materials constituting the main shaft 107 and the main bearing 111 and the design specifications, in the present embodiment, the rigidity ratio Gb between the main shaft 107 and the main bearing 111 in the straight portion 123 of the thinnest main bearing 111. / Gc is 1.85, which is the minimum value in the length direction of the main bearing 111. The ratio L / D between the length L and the inner diameter D of the main bearing 111 is 2.20.

  About the compressor comprised as mentioned above, the operation | movement is demonstrated below.

  When the electric element 105 is energized, the rotor 104 rotates and the crankshaft 109 rotates accordingly. The main shaft 107 rotates while being supported by the main bearing 111, while the rotational motion of the eccentric shaft 108 is transmitted to the piston 114 through the connecting means 117 and the piston pin 115, and the piston 114 reciprocates in the bore hole 113. To do. Due to the reciprocating motion of the piston 114, the refrigerant 121 in the sealed container 101 is compressed after being sucked into the compression chamber 112 and discharged outside the sealed container 101. The so-called suction stroke and compression stroke are repeated.

  The oil 102 is sucked into the oil pump hole 119 by the centrifugal force accompanying the rotation of the crankshaft 109, guided upward from the spiral groove 118 through the communication hole 120, and injected from the upper end of the eccentric shaft 108. However, the sliding means such as the connecting means 117 and the piston pin 115 and the piston 114 and the bore hole 113 are lubricated.

  The behavior of the main shaft 107 during the compression stroke will be described with reference to FIG.

  When the compressive load P acts on the eccentric shaft 108 from the right side via the connecting means 117, the main shaft 107 is tilted to the left, above the main bearing 111 (point U in FIG. 3) and below (FIG. 3). Middle, point M). However, in the present embodiment, the rigidity ratio Gb / Gc between the bending rigidity Gb of the main bearing 111 and the bending rigidity Gc of the main shaft 107 in the straight portion 123 that is the thinnest in the main bearing 111 is 1.8 or more. As a result, the bending rigidity of the main shaft 107 is significantly smaller than that of the main bearing 111, and the main shaft 107 is more easily elastically deformed than the main bearing 111.

  As a result, when the main shaft 107 comes into contact with the main bearing 111, the main shaft 107 is bent starting from a portion that contacts the lower portion of the main bearing 111 (point M in FIG. 3). More specifically, tensile stress is applied to the right side (the side in contact with the point M in FIG. 3) of the sliding surface of the main spindle 107, and compressive stress is applied to the left side (the side in contact with the point U in FIG. 3). In this way, the main shaft 107 is curved as shown in FIG. Therefore, the lower portion of the main shaft 107 follows the inner diameter shape of the lower portion of the main bearing 111 (point M in FIG. 3) over a relatively wide range, and the contact area between the main shaft 107 and the main bearing 111 is increased, resulting in locality. Avoids contact sliding and reduces wear by one piece.

  Here, with reference to FIG. 4, the amount of wear at the main shaft 107 portion that contacts the upper side (point U in FIG. 3) and the lower side (point M in FIG. 3) of the main bearing 111 in the present embodiment will be described. To do.

  In this figure, the present embodiment in which the rigidity ratio Gb / Gc is 1.85 is compared with the conventional example in which the rigidity ratio Gb / Gc is 1.65. In any form, the ratio L / D between the length L of the main bearing 111 and the inner diameter D was 2.2. In this study, the compressor is continuously operated for about 20 days at the maximum condensation / evaporation pressure value in the actual working pressure range of the refrigerator in a short circuit cycle.

  From FIG. 4, it was confirmed by an actual machine test that the amount of wear at the main shaft 107 part contacting the lower part of the main bearing 111 (point M in FIG. 3) was reduced by about 50% compared to the conventional example.

  This is because, by bending the main shaft 107, the contact area between the main shaft 107 and the main bearing 111 increases along the inner diameter shape below the main bearing 111 (point M in FIG. 3) over a relatively wide range. In addition to avoiding local contact sliding, it is considered that the lubricating action by the oil 102 was also exhibited synergistically.

  Next, the rigidity ratio Gb / Gc between the bending rigidity Gb below the main bearing 111 and the bending rigidity Gc of the main shaft 107 and the amount of wear measured at the main shaft portion in contact with the main bearing (point M in FIG. 3). The correlation will be described with reference to FIG.

  In this examination, the ratio L / D between the length L of the main bearing 111 and the inner diameter D was set to 2.20, and a 6-spec compressor was produced using the rigidity ratio Gb / Gc as a parameter. These compressors are continuously operated for about 20 days in a short circuit cycle at a condensing / evaporation pressure value that is maximum in the actual working pressure range of the refrigerator.

  From FIG. 5, when the rigidity ratio Gb / Gc is 1.8 or more, the wear amount of the main shaft 107 is 1 μm or less, and it is confirmed that the wear is reduced due to one piece. On the other hand, when the rigidity ratio Gb / Gc was between 1.4 and 1.8, it was confirmed that the main shaft 107 was worn by one piece.

  In addition, when the rigidity ratio Gb / Gc is 1.4, it is confirmed that although the wear due to the contact of the main shaft 107 is suppressed to about 1 μm, the vibration of the compressor during operation becomes excessively large.

  This is because the main bearing 111 has a significantly smaller bending rigidity than the main shaft 107 and is easily elastically deformed, so that the lower side of the main bearing 111 has a larger inclination angle of the main shaft 107 due to the pressing force due to the inclination of the main shaft 107. This is considered to be a result of being excessively curved in the direction of the movement.

  Therefore, although the local contact sliding is avoided because the inner diameter of the main bearing 111 is along the main shaft 107, the tilt angle of the main shaft 107 becomes larger, so that the swing of the main shaft 107 during rotation becomes excessive. This is thought to have led to an increase in compressor vibration.

  From this, if the rigidity ratio Gb / Gc is 1.8 or more, it not only reduces wear caused by one-piece contact between the main shaft 107 and the main bearing 111, but also suppresses vibration of the compressor. Is also possible.

  Next, regarding the correlation between the ratio L / D of the main bearing 111 length L and the inner diameter D and the measured amount of wear at the main shaft 107 part in contact with the lower side of the main bearing 111 (point M in FIG. 3), This will be described with reference to FIG.

  In this figure, based on the present embodiment in which the rigidity ratio Gb / Gc is 1.85 and the conventional example in which the rigidity ratio Gb / Gc is 1.65, compressors of 6 specifications each with the ratio L / D as a parameter. Was made. These compressors are continuously operated for about 20 days in a short circuit cycle at a condensing / evaporation pressure value that is maximum in the actual working pressure range of the refrigerator.

  From FIG. 6, in the region where the ratio L / D is 2.5 or less, it was confirmed that the effect of reducing wear due to one piece was significantly reduced by setting the rigidity ratio Gb / Gc to 1.8 or more.

  According to the present embodiment, it is possible to reduce the length of the main shaft 107 and the main bearing 111 while ensuring sufficient reliability, and to meet the recent demand for downsizing of the compressor.

  In the case where the ratio L / D is larger than 2.5, the following is considered.

  The longer the length L of the main bearing 111, the smaller the contact load between the main shaft 107 and the main bearing 111 due to the moment relationship. it is conceivable that. In addition, the smaller the inner diameter D of the main bearing 111, that is, the outer diameter of the main shaft 107, the smaller the bending rigidity and the more easily the main shaft 107 bends. It is thought that.

  From the above, by setting the rigidity ratio Gb / Gc between the bending rigidity Gb of the main bearing 111 and the bending rigidity Gc below the main shaft 107 to 1.8 or more as in the present embodiment, the bending rigidity of the main shaft 107 is increased. Is significantly lowered with respect to the main bearing 111 so that when the main shaft 107 is inclined and comes into contact with the inside of the main bearing 111 during the compression stroke, the main shaft 107 contacts a portion below the main bearing 111. Since it curves to the starting point and follows the inner diameter shape of the main bearing 111, the contact area between the main shaft 107 and the main bearing 111 increases, wear due to one piece decreases, and a highly reliable and inexpensive compressor can be provided. .

  In addition, in the region where the ratio L / D between the length L of the main bearing 111 and the inner diameter D is 2.5 or less, the effect of reducing wear due to the contact of the main shaft 107 is noticeable by designating the range of the rigidity ratio. In addition, the length of the main bearing 111 and the main shaft 107 can be shortened to reduce the size of the compressor.

  In the present embodiment, the main bearing 111 is formed of a material higher than the Young's modulus of the material of the main shaft 107, and the productivity is excellent without significantly changing the current production line.

(Embodiment 2)
FIG. 7 is a longitudinal sectional view of a compressor according to Embodiment 2 of the present invention, and FIG. 9 is a characteristic diagram showing the behavior of the main shaft in the same embodiment, FIG. 10 is a characteristic diagram showing a measurement result of the main shaft wear amount in the same embodiment, and FIG. 11 is a diagram showing the rigidity ratio Gb / Gc. FIG. 12 is a characteristic diagram showing the correlation result between the main bearing length / inner diameter ratio L / D and the main shaft wear amount.

  The hermetic container 201 is filled with isobutane (R600a) as the refrigerant 221 and stores mineral oil having a relatively low viscosity as the oil 202.

  The electric element 205 includes a stator 203 fixed below the cylinder block 210 and connected to an inverter drive circuit (not shown), and a rotor 204 containing a permanent magnet and fixed below the main shaft 207. An electric motor for driving the inverter is formed, and is driven at a plurality of operation frequencies including an operation frequency (for example, 1500 r / min) lower than the commercial power supply frequency by the inverter drive circuit.

  The stator 203 is a salient pole concentrated winding type winding. The salient pole concentrated winding type winding has an iron core having a portion (a salient pole) projecting radially toward the center of the rotor 204, and a covered wire is concentrated around the salient pole of the iron core. It is attached. Since the salient pole concentrated winding has very little protrusion from the iron core, the electric element 205 has a lower overall height than other winding types (for example, distributed winding type).

  The compression element 206 will be described below.

  The crankshaft 209 extends in the vertical direction and includes a main shaft 207 and an eccentric shaft 208. The cylinder block 210 has a bore 213 that forms a compression chamber 212 together with a main bearing 211 that rotatably supports the main shaft 207.

  The piston 214 is reciprocally inserted into the bore hole 213. The piston pin 215 has a substantially cylindrical shape, is disposed in parallel with the eccentric shaft 208, and is non-rotatably locked in a piston pin hole 216 formed in the piston 214. The connecting means 217 connects the eccentric shaft 208 and the piston 214 via the piston pin 215.

  The inner diameter D of the main bearing 211 is set to 18 mm, which can effectively suck up the oil 202 by applying centrifugal force effectively even at low speed operation (for example, the operation frequency is 1500 r / min).

  The clearance between the outer diameter of the main shaft 207 and the inner diameter D of the main bearing 211 is within a range of 10 to 30 μm in consideration of machining accuracy, twisting due to a change in diameter during heat, or prevention of rattling during rotation of the main shaft 207. It is. The length L of the main bearing 211 is 40 mm.

  The main shaft 207 is connected to the spiral groove 218 formed on the sliding surface with the main bearing 211, the lower end of the spiral groove 218 through the communication hole 220, and the oil pump hole 219 that opens to the oil 202. A cylindrical hole 230 extending at the upper end of the pump hole 219 is provided.

  The oil pump hole 219 is formed so as to incline upward from the lower end opened to the oil 202. The inner diameter of the oil pump hole 219 is 11.5 mm. The cylindrical hole 230 is inclined at the same inclination angle as that of the oil pump hole 219, and is drilled to a position slightly above the middle of the sliding portion of the main shaft 207 and the main bearing 211. The hole diameter of the cylindrical hole 230 is 10 mm.

  Both the crankshaft 209 and the cylinder block 210 are made of gray cast iron (FC200), which is an inexpensive and easy to process cast iron material.

  Due to the structure of the compressor, the straight portion 223 of the main bearing 211 surrounded by the recess 222 formed on the upper end surface of the rotor 204 corresponds to the thinnest portion. The minimum thickness below the main bearing 211 is 2.5 mm in consideration of processing, assembly tolerance, impact cracking, brittle cracking due to metal fatigue during long-term operation, and the like. Therefore, the outer diameter of the straight portion 223 of the main bearing 211 is 23 mm.

  Here, the bending rigidity indicates the magnitude of deformation resistance to bending, and is represented by the product of Young's modulus and cross-sectional second moment.

  In the present embodiment, from the design and assembly specifications that constitute the main shaft 207 and the main bearing 211, in the present embodiment, the straight portion 223 of the main bearing 211 that is the thinnest and the main shaft 207 portion in which the cylindrical hole 230 is formed. The rigidity ratio Gb / Gc between the main shaft 207 and the main bearing 211 is 1.85, which is the minimum value in the length direction of the main bearing 211. The ratio L / D between the length L and the inner diameter D of the main bearing 211 is 2.20.

  About the compressor comprised as mentioned above, the operation | movement is demonstrated below.

  When the electric element 205 is energized, the rotor 204 rotates and the crankshaft 209 also rotates accordingly. The main shaft 207 rotates while being supported by the main bearing 211, while the rotational motion of the eccentric shaft 208 is transmitted to the piston 214 through the connecting means 217 and the piston pin 215, and the piston 214 reciprocates in the bore hole 213. To do. Due to the reciprocating motion of the piston 214, the refrigerant 221 in the sealed container 201 is sucked into the compression chamber 212 and then compressed and discharged outside the sealed container 201. The so-called suction stroke and compression stroke are repeated.

  The oil 202 is sucked into the oil pump hole 219 by the centrifugal force accompanying the rotation of the crankshaft 209, guided upward from the spiral groove 218 through the communication hole 220, and injected from the upper end of the eccentric shaft 208. However, the sliding means such as the connecting means 217 and the piston pin 215 and the piston 214 and the bore hole 213 are lubricated.

  The behavior of the main shaft 207 during the compression stroke will be described with reference to FIG.

  When the compressive load P acts on the eccentric shaft 208 from the right side via the connecting means 217, the main shaft 207 tilts to the left, above the main bearing 211 (point U in FIG. 9) and below (FIG. 9). Middle, point M). However, in the present embodiment, the rigidity ratio Gb / Gc between the bending rigidity Gb of the main bearing 211 and the bending rigidity Gc of the main shaft 207 in the straight portion 223 that is the thinnest in the main bearing 211 is 1.8 or more. Thus, the bending rigidity of the main shaft 207 is significantly smaller than that of the main bearing 211, and the main shaft 207 is more easily elastically deformed than the main bearing 211.

  As a result, when the main shaft 207 comes into contact with the main bearing 211, the main shaft 207 is curved starting from a portion that contacts the lower portion of the main bearing 211 (point M in FIG. 9). More specifically, tensile stress is applied to the right side (the side in contact with the point M in FIG. 9) of the sliding surface of the main spindle 107, and compressive stress is applied to the left side (the side in contact with the point U in FIG. 9). In this way, the main shaft 207 is curved as shown in FIG. Therefore, the lower part of the main shaft 207 follows the inner diameter shape of the lower part of the main bearing 211 (point M in FIG. 9) over a relatively wide range, the contact area between the main shaft 207 and the main bearing 211 increases, Avoids contact sliding and reduces wear by one piece.

  Here, the amount of wear at the main shaft 207 portion contacting the upper side (point U in FIG. 9) and the lower side (point M in FIG. 9) of the main bearing 211 in the present embodiment will be described with reference to FIG. To do.

  In this figure, the present embodiment in which the rigidity ratio Gb / Gc is 1.85 is compared with the conventional example in which the rigidity ratio Gb / Gc is 1.65. In any form, the ratio L / D between the length L of the main bearing 211 and the inner diameter D was 2.2. In this study, the compressor is continuously operated for about 20 days at the maximum condensation / evaporation pressure value in the actual working pressure range of the refrigerator in a short circuit cycle.

  From FIG. 10, it was confirmed by an actual machine test that the amount of wear at the main shaft 207 portion in contact with the lower part of the main bearing 211 (point M in FIG. 9) was reduced by about 60% compared to the conventional example.

  This is because the main shaft 207 is curved, so that the contact area between the main shaft 207 and the main bearing 211 increases along a relatively wide inner diameter shape below the main bearing 211 (point M in FIG. 9). In addition to avoiding local contact sliding, it is considered that the lubricating action by the oil 202 was also synergistically exhibited.

  Next, the rigidity ratio Gb / Gc between the bending rigidity Gb below the main bearing 211 and the bending rigidity Gc of the main shaft 207 and the amount of wear measured at the main shaft portion in contact with the main bearing (point M in FIG. 9). The correlation will be described with reference to FIG.

  In this examination, the ratio L / D between the length L of the main bearing 211 and the inner diameter D was set to 2.20, and a 6-spec compressor was produced using the rigidity ratio Gb / Gc as a parameter. These compressors are continuously operated for about 20 days in a short circuit cycle at a condensing / evaporation pressure value that is maximum in the actual working pressure range of the refrigerator.

  From FIG. 11, when the rigidity ratio Gb / Gc is 1.8 or more, the wear amount of the main shaft 207 is 1 μm or less, and it is confirmed that the wear is reduced due to one piece. On the other hand, when the rigidity ratio Gb / Gc was between 1.4 and 1.8, it was confirmed that the main shaft 207 was worn by one piece.

  Further, when the rigidity ratio Gb / Gc is 1.4, it is confirmed that the vibration due to the contact of the main shaft 207 with one piece is suppressed to about 1 μm, but the vibration of the compressor during operation becomes excessively large.

  This is because the main bearing 211 has a significantly smaller bending rigidity than the main shaft 207 and is easily elastically deformed. The pressing force caused by the inclination of the main shaft 207 causes the lower side of the main bearing 211 to increase the inclination angle of the main shaft 207. This is considered to be a result of being excessively curved in the direction of the movement.

  Therefore, although the local contact sliding is avoided by having the inner diameter of the main bearing 211 along the main shaft 207, the tilt angle of the main shaft 207 becomes larger, so that the whirling during rotation of the main shaft 207 becomes excessive. This is thought to have led to an increase in compressor vibration.

  From this, if the rigidity ratio Gb / Gc is 1.8 or more, it not only reduces wear caused by one-piece contact between the main shaft 207 and the main bearing 211, but also suppresses vibration of the compressor. Is also possible.

  Next, regarding the correlation between the ratio L / D of the main bearing 211 length L and the inner diameter D and the measured amount of wear at the main shaft 207 portion in contact with the lower side of the main bearing 211 (point M in FIG. 9), This will be described with reference to FIG.

  In this figure, based on the present embodiment in which the rigidity ratio Gb / Gc is 1.85 and the conventional example in which the rigidity ratio Gb / Gc is 1.65, compressors of 6 specifications each with the ratio L / D as a parameter. Was made. These compressors are continuously operated for about 20 days in a short circuit cycle at a condensing / evaporation pressure value that is maximum in the actual working pressure range of the refrigerator.

  From FIG. 12, in the region where the ratio L / D is 2.5 or less, it was confirmed that the effect of reducing wear due to one piece was significantly reduced by setting the rigidity ratio Gb / Gc to 1.8 or more.

  According to the present embodiment, it is possible to reduce the lengths of the main shaft 207 and the main bearing 211 while ensuring sufficient reliability and meet the recent demand for downsizing of the compressor.

  In the case where the ratio L / D is larger than 2.5, the following is considered.

  The longer the length L of the main bearing 211, the smaller the contact load between the main shaft 207 and the main bearing 211 due to the moment relationship. For example, even the local contact sliding does not reach the surface pressure level at which wear proceeds. it is conceivable that. Further, the smaller the inner diameter D of the main bearing 211, that is, the outer diameter of the main shaft 207, the smaller the bending rigidity and the more easily the main shaft 207 bends. It is thought that.

  From the above, by setting the rigidity ratio Gb / Gc between the bending rigidity Gb of the main bearing 211 and the bending rigidity Gc below the main shaft 207 to 1.8 or more as in the present embodiment, the bending rigidity of the main shaft 207 is increased. Is significantly lowered with respect to the main bearing 211, so that when the main shaft 207 is inclined and contacts the inside of the main bearing 211 when receiving the compressive load P during the compression stroke, the main shaft 207 makes a contact with the lower portion of the main bearing 211. Since it curves to the starting point and follows the inner diameter shape of the main bearing 211, the contact area between the main shaft 207 and the main bearing 211 increases, wear due to one piece decreases, and a highly reliable and inexpensive compressor can be provided. .

  Further, as in the present embodiment, by extending the cylindrical hole 230 at the upper end of the oil pump hole 219, the bending rigidity of the spindle can be easily reduced only by drilling, and the productivity is excellent. Yes. In addition, since it corresponds by changing the hole diameter and length of the cylindrical hole 230 extended at the upper end of the oil pump hole 219 according to the compression load condition required for the compressor, it is extremely excellent in versatility. .

  In this embodiment, both the main bearing 211 and the main shaft 207 are made of the same type of cast iron material. However, in order to set the rigidity ratio Gb / Gc to 1.8 or less, in addition to processing the cylindrical hole, The bearing 211 may be formed of a material higher than the Young's modulus of the material of the main shaft 207.

  As described above, since the compressor according to the present invention has high reliability, it can be applied to all uses using a refrigeration cycle such as a dehumidifier, a showcase, and a vending machine, including a refrigerator for home use.

The longitudinal cross-sectional view of the compressor in Embodiment 1 of this invention Sectional drawing in the A-A 'part of the main axis | shaft and main bearing in Embodiment 1 of this invention FIG. 3 is a characteristic diagram showing the behavior of the main shaft according to the first embodiment of the present invention. The characteristic view which shows the measurement result of the spindle wear amount of Embodiment 1 of this invention The characteristic view which shows the correlation result of rigidity ratio Gb / Gc and spindle wear amount of Embodiment 1 of this invention The characteristic view which shows the correlation result of ratio L / D of Embodiment 1 of this invention, and spindle wear amount The longitudinal cross-sectional view of the compressor in Embodiment 2 of this invention Sectional drawing in the B-B 'part of the main axis | shaft and main bearing in Embodiment 2 of this invention The characteristic figure which shows the behavior of the main axis | shaft of Embodiment 2 of this invention The characteristic view which shows the measurement result of the spindle wear amount of Embodiment 2 of this invention The characteristic view which shows the correlation result of rigidity ratio Gb / Gc and spindle wear amount of Embodiment 2 of this invention The characteristic view which shows the correlation result of ratio L / D and spindle wear amount of Embodiment 2 of this invention Vertical section of a conventional compressor Characteristic diagram showing the behavior of the main shaft of a conventional compressor

Explanation of symbols

101, 201 Airtight container 102, 202 Oil 103, 203 Stator 104, 204 Rotor 105, 205 Electric element 106, 206 Compression element 107, 207 Main shaft 109, 209 Crankshaft 110, 210 Cylinder block 111, 211 Main bearing 118, 218 Spiral groove 119, 219 Oil pump hole 123, 223 Straight part 230 Cylindrical hole

Claims (5)

Oil is stored in a sealed container, and an electric element composed of a stator and a rotor, and a compression element driven by the electric element are accommodated. The compression element extends in a vertical direction and has a main shaft. A cylinder block having a shaft and a main bearing that supports the main shaft, and a rigidity ratio Gb between the bending rigidity Gb of the main bearing and the bending rigidity Gc of the main shaft in the thinnest straight portion of the main bearing; / the Gc, compressor was 4.6 or less 1.8 or more. The compressor according to claim 1, wherein a ratio L / D between the main bearing length L and the main bearing inner diameter D is 1.9 or more and 2.5 or less . The compressor according to claim 1 or 2, wherein the material of the main bearing is spheroidal graphite cast iron or gray cast iron, and the material of the main shaft is gray cast iron . The main shaft includes a spiral groove formed on a sliding surface with the main bearing, an oil pump hole that communicates with a lower end of the spiral groove and opens to oil, and a cylindrical hole extends to the upper end of the oil pump hole. The compressor according to any one of claims 1 to 3, wherein the compressor is drilled. The compressor according to any one of claims 1 to 4, wherein the compressor is operated at an operation frequency including at least a frequency lower than a commercial power supply frequency.

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