WO2024071128A1

WO2024071128A1 - Hermetic refrigerant compressor, operating method for same, and freezing/refrigerating device using same

Info

Publication number: WO2024071128A1
Application number: PCT/JP2023/034966
Authority: WO
Inventors: 政信権藤; 健汰森
Original assignee: パナソニックアプライアンシズリフリジレーションデヴァイシズシンガポール; パナソニックＩｐマネジメント株式会社
Priority date: 2022-09-30
Filing date: 2023-09-26
Publication date: 2024-04-04

Abstract

This hermetic refrigerant compressor comprises: a hermetic container; refrigerator oil that is stored inside the hermetic container and has a kinematic viscosity at 40°C that is within the range of 1.0 mm²/s to 2.5 mm²/s; a cylinder block that is accommodated inside the hermetic container and forms a compression chamber; and a piston inserted into the compression chamber so as to be capable of reciprocating. When the operating frequency is 16-35 r/s, the average speed at which the piston reciprocates is set to exceed 0.31 m/s.

Description

Hermetic refrigerant compressor, its operating method, and refrigeration/freezing device using the same

The present invention relates to a hermetic refrigerant compressor used in refrigerators, air conditioners, etc., a method for operating the hermetic refrigerant compressor, and a freezing/refrigeration device using the hermetic refrigerant compressor.

In recent years, in order to protect the global environment, there has been progress in the development of highly efficient hermetic refrigerant compressors that require less input power and therefore less use of fossil fuels. Reducing the operating frequency is an effective way to reduce the input power of hermetic refrigerant compressors.

On the other hand, when it comes to high efficiency, the coefficient of performance (COP), which is expressed as refrigeration capacity/input power, is an index that indicates the efficiency of a hermetic refrigerant compressor. In order to improve the coefficient of performance (COP), for example, it has been proposed to use oil with a lower viscosity (refrigeration oil, lubricant oil). For example, Patent Document 1 discloses a refrigerant compressor that achieves high efficiency by setting the viscosity of the oil stored inside the hermetic container to VG3 or more and VG8 or less.

JP 2008-531896 A

In recent years, efforts have been made to further reduce the viscosity of refrigeration oil in order to further improve the coefficient of performance (COP). For example, in the refrigerant compressor disclosed in Reference 1, the oil (refrigeration oil) viscosity is set to be in the range of VG3 to VG8 as mentioned above, but in recent years, the use of refrigeration oil of VG3 or less has also been considered. However, simply using a lower viscosity refrigeration oil to further reduce the sliding loss is not sufficient to further improve the coefficient of performance (COP).

The present invention has been made to solve these problems, and aims to provide a hermetic refrigerant compressor and an operating method thereof that can achieve an even better coefficient of performance (COP) by using a lower viscosity refrigeration oil, as well as a freezing/refrigeration device using the same.

In order to solve the above-mentioned problems, the hermetic refrigerant compressor of the present disclosure comprises a hermetic container, refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to ^{2.5 mm2} /s, a cylinder block housed in the hermetic container and forming a compression chamber, and a piston inserted and reciprocatable inside the compression chamber, wherein when the operating frequency is 16 r/s or more and 35 r/s or less, the average reciprocating speed of the piston is set to exceed 0.31 m/s.

According to the above configuration, when the operating frequency of the hermetic refrigerant compressor is 16 r/s or more and 35 r/s or less, the lower limit of the average speed of the reciprocating movement of the piston (average piston speed) is set. As a result, even if the operating speed of the hermetic refrigerant compressor is relatively low, the average piston speed can be relatively high. Therefore, even if a low-viscosity oil with a further reduced viscosity (kinematic viscosity at 40°C set within the range of 1.0 mm ² /s to 2.5 mm ² /s) is used as the refrigerating machine oil, a good oil film can be formed between the piston reciprocating at high speed and the compression chamber. As a result, it is possible to effectively suppress leakage of refrigerant gas from the piston and the compression chamber. Therefore, when a low-viscosity oil is used as the refrigerating machine oil, the coefficient of performance (COP) of the hermetic refrigerant compressor can be further improved.

In order to solve the above-mentioned problems, the hermetic refrigerant compressor according to the present disclosure may include a hermetic container, refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to ^{2.5 mm2} /s, a cylinder block housed in the hermetic container and forming a compression chamber, and a piston inserted and reciprocatable within the compression chamber, wherein a ratio S/D of a reciprocating stroke amount (S) of the piston to a piston diameter (D) is in the range of 0.78 to 1.00.

According to the above configuration, when a low-viscosity oil having a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 mm ² /s is used as the refrigeration oil, the ratio S/D is set within a predetermined range. This allows the stroke amount (S) to be relatively large, and therefore the average piston speed to be relatively large. As a result, the amount of refrigerant gas leakage can be effectively suppressed.

In addition, when the low viscosity oil is used as the refrigeration oil, the piston diameter (D) can be made relatively small by setting the ratio S/D within a predetermined range. This makes it possible to reduce the total area of the clearance between the outer peripheral surface of the piston and the inner peripheral surface of the cylinder, thereby effectively suppressing leakage of refrigerant gas.

Furthermore, if the piston diameter (D) is relatively small, the compression load of the refrigerant gas on the piston can be reduced. This allows the input power required to reciprocate the piston to be reduced.

Therefore, by setting the above-mentioned ratio S/D within a specified range, it is possible to further improve the coefficient of performance (COP) of the hermetic refrigerant compressor when using low-viscosity oil as refrigeration oil.

In order to solve the above-mentioned problems, the hermetic type refrigerant compressor according to the present disclosure may include a hermetic container, refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to 2.5 ^mm2 /s, a cylinder block contained in the hermetic container and forming a compression chamber, and a piston inserted reciprocally within the compression chamber, wherein a ratio L1/D of the piston overall length (L1) to the piston diameter (D) is in the range of 0.8 to 1.0, and when the length of an area where the piston seals the compression chamber by its reciprocating motion is defined as a sealing length (L2), a ratio L2/L1 of the sealing length (L2) to the piston overall length (L1) is in the range of 0.9 to 1.0.

According to the above configuration, when a low-viscosity oil having a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 mm ² /s is used as the refrigeration oil, the ratios L1/D and L2/L1 are set within a predetermined range. This makes it possible to relatively increase the seal length (L2) without excessively increasing the overall piston length (L1) while suppressing an increase in sliding loss. As a result, it is possible to increase the viscous force of the oil film between the outer circumferential surface of the piston and the inner circumferential surface of the cylinder while suppressing an increase in sliding loss.

In addition, the area between the piston and cylinder that is sealed by the oil film of the refrigeration oil can be made relatively large. This makes it possible to further suppress leakage of refrigerant gas.

Furthermore, if the seal length (L2) is increased, it becomes possible to stabilize the position of the piston reciprocating in the compression chamber. As a result, it becomes possible to further suppress the increase in sliding loss.

Therefore, by setting the ratio L1/D and the ratio L2/L1 within a specified range, the coefficient of performance (COP) of the hermetic refrigerant compressor can be further improved when low-viscosity oil is used as the refrigeration oil.

In addition, in order to solve the above-mentioned problems, the operating method of a hermetic refrigerant compressor according to the present disclosure comprises a hermetic container, refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2/ s to 2.5 ^mm2 /s, a cylinder block contained in the hermetic container and forming a compression chamber, and a piston inserted and capable of reciprocating within the compression chamber, wherein when the operating frequency is 16 r/s or more and 35 r/s or less, the average reciprocating speed of the piston exceeds 0.31 m/s.

Furthermore, the freezing/refrigeration device according to the present disclosure may be configured to include a hermetic refrigerant compressor of the above-described configuration, a radiator, a pressure reducing device, and a heat sink, and to have a refrigerant circuit in which these are connected in a ring shape by piping.

The above objects, other objects, features, and advantages of the present invention will become apparent from the following detailed description of the preferred embodiments, which are to be read in conjunction with the accompanying drawings.

The present invention has the effect of providing a hermetic refrigerant compressor and an operating method thereof that can achieve an even better coefficient of performance (COP) by using a refrigeration oil with a lower viscosity, as well as a freezing/refrigeration device using the same.

FIG. 1 is a schematic cross-sectional view illustrating an example of a configuration of a hermetic refrigerant compressor according to an embodiment of the present disclosure. FIG. 2 is a schematic enlarged side view of a piston and a cylinder in a compression element included in the hermetic refrigerant compressor shown in FIG. 1 . FIG. 3A is a schematic side view showing an example of the configuration of a piston included in the hermetic refrigerant compressor shown in FIG. 1, and FIG. 3B is a schematic side view showing an example of the configuration of a conventional piston. FIG. 4A is a schematic diagram showing an example of a configuration of a crankshaft included in the hermetic refrigerant compressor shown in FIG. 1 when the sliding surface is a single surface, and FIGS. 4B and 4C are schematic diagrams showing an example of a configuration of the crankshaft shown in FIG. 3A when the sliding surface is divided into a plurality of surfaces. FIG. 2 is a partial cross-sectional view showing an example of a distance P and a distance Q in the hermetic refrigerant compressor shown in FIG. 1 , and a load (main shaft load) applied to a main shaft sliding portion. FIG. 2 is a partial cross-sectional view showing an example of a configuration of a main part of a thrust bearing in the hermetic refrigerant compressor shown in FIG. 1 . FIG. 2 is a schematic diagram showing an example of the configuration of a refrigeration/freezing device including the hermetic refrigerant compressor shown in FIG. 1 . 1 is a graph showing a relationship (characteristics) of the amount of refrigerant gas leakage versus the kinetic viscosity of refrigeration oil when the operating frequency is 17 r/s in a conventional refrigerant compressor, as a reference example of the present disclosure. 1 is a graph showing a relationship (characteristic) of the coefficient of performance (COP) with respect to the kinetic viscosity of a refrigeration oil when the operating frequency is 27 r/s, in an embodiment of the present disclosure and a conventional example. 1 is a graph showing a relationship (characteristic) of the coefficient of performance (COP) with respect to the kinetic viscosity of a refrigeration oil when the operating frequency is 17 r/s, in an embodiment of the present disclosure and a conventional example.

(Knowledge and other information that forms the basis of this disclosure)
In a hermetic refrigerant compressor, a compression chamber is formed in a cylinder of a cylinder block, and a piston is inserted in the compression chamber so as to be able to reciprocate. Refrigeration oil exists as an oil film between the outer peripheral surface of the piston and the inner peripheral surface of the cylinder (the inner peripheral surface of the compression chamber), and lubricates the reciprocating motion (i.e., sliding) of the piston. If the space between the outer peripheral surface of the piston and the inner peripheral surface of the cylinder is abbreviated as "piston-cylinder space" for convenience of explanation, the oil film formed between the piston and cylinder becomes thinner if the kinetic viscosity of the refrigeration oil becomes smaller. The thinner oil film makes it possible to reduce the sliding loss accompanying the reciprocating motion (sliding) of the piston, but it is also assumed that refrigerant gas may leak from between the piston and cylinder.

As will be described in the Examples below, the inventors experimentally verified and evaluated the amount of refrigerant gas leaking from between the piston and cylinder in a hermetic refrigerant compressor having a conventional configuration, using refrigerating machine oils with different kinetic viscosities at 40° C. under conditions of a low operating frequency of, for example, 17 ^{r/s (rps).As a result, it was revealed that the amount of refrigerant gas leaking increases when the kinetic viscosity of the refrigerating machine oil at 40° C. is 2.5 mm 2} /s or less.

Based on this verification and evaluation, the amount of refrigerant gas leakage tends to increase as the kinetic viscosity at 40° C. of the refrigerating machine oil used decreases. However, even if the viscosity of the refrigerating machine oil is reduced up to 2.5 ^mm2 /s, it is considered that the reduction in friction loss in the hermetic refrigerant compressor can be more than offset by the effect of the increase in the amount of refrigerant gas leakage, and as a result, the coefficient of performance (COP) can be improved.

However, if the kinetic viscosity of the refrigeration oil at 40° C. is 2.5 mm ² /s or less, the amount of leakage of the refrigerant gas becomes too large, and the refrigeration capacity of the refrigeration cycle equipped with a hermetic refrigerant compressor decreases, and as a result, it has become clear that the coefficient of performance (COP) cannot be improved.

If it becomes possible to effectively suppress the leakage of refrigerant gas from between the piston and cylinder while reducing the kinetic viscosity of the refrigeration oil, it is expected that the coefficient of performance (COP) of hermetic refrigerant compressors will be further improved. However, it is difficult to simultaneously reduce the viscosity of the refrigeration oil and suppress the amount of refrigerant gas leakage.

As a result of further intensive research, the inventors discovered that it is possible to improve the coefficient of performance (COP) by effectively preventing refrigerant gas leakage from between the piston and cylinder, which led to the completion of this invention.

In other words, the hermetic refrigerant compressor of the present disclosure comprises a hermetic container, refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to ^{2.5 mm2} /s, a cylinder block housed in the hermetic container and forming a compression chamber, and a piston inserted reciprocally within the compression chamber, and when the operating frequency is 16 r/s or more and 35 r/s or less, the average reciprocating speed of the piston is set to greater than 0.31 m/s.

In a hermetic refrigerant compressor having the above configuration, the ratio S/D of the reciprocating stroke amount (S) of the piston to the piston diameter (D) may be within the range of 0.78 to 1.00.

In this configuration, in addition to relatively increasing the average piston speed, the ratio S/D is set within the above range. By setting this ratio S/D, it is possible to relatively increase the stroke amount (S) and relatively decrease the piston diameter (D), as described below.

This effectively suppresses the amount of refrigerant gas leakage and reduces the input power required to reciprocate the piston. Therefore, when low-viscosity oil is used as the refrigeration oil, the coefficient of performance (COP) of the hermetic refrigerant compressor can be further improved.

Furthermore, in the hermetic refrigerant compressor of the above configuration, when the length of the area where the piston seals the inside of the compression chamber by its reciprocating motion is defined as the sealing length (L2), the ratio L1/D of the piston overall length (L1) to the piston diameter (D) may be within a range of 0.8 to 1.0, and the ratio L2/L1 of the sealing length (L2) to the piston overall length (L1) may be within a range of 0.9 to 1.0.

In this configuration, in addition to relatively increasing the average piston speed, the ratios L1/D and L2/L1 are set within a predetermined range. By setting these ratios, it is possible to relatively increase the seal length (L2) without excessively increasing the overall piston length (L1), as described below.

This makes it possible to increase the viscosity of the oil film between the outer circumferential surface of the piston and the inner circumferential surface of the cylinder while suppressing an increase in sliding loss, and also to further suppress the leakage of refrigerant gas. As a result, the coefficient of performance (COP) of the hermetic refrigerant compressor can be further improved.

Another hermetic refrigerant compressor according to the present disclosure comprises a hermetic container, refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to ^{2.5 mm2} /s, a cylinder block housed in the hermetic container and forming a compression chamber, and a piston inserted reciprocally within the compression chamber, wherein the ratio S/D of the reciprocating stroke amount (S) of the piston to the piston diameter (D) is in the range of 0.78 to 1.00.

In a hermetic refrigerant compressor having the above configuration, when the length of the area where the piston seals the inside of the compression chamber by its reciprocating motion is defined as a sealing length (L2), the ratio L2/L1 of the sealing length (L2) to the total piston length (L1) may be within the range of 0.9 to 1.0.

According to the above configuration, in addition to setting the ratio S/D of the piston stroke amount (S) to the piston diameter (D) within a predetermined range, the ratios L1/D and L2/L1 are also set within predetermined ranges. By setting these ratios, as described below, it is possible to relatively increase the seal length (L2) without excessively increasing the overall piston length (L1).

Another hermetic refrigerant compressor according to the present disclosure comprises a hermetic container, refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to ^{2.5 mm2} /s, a cylinder block housed in the hermetic container and forming a compression chamber, and a piston inserted reciprocally within the compression chamber, wherein a ratio L1/D of the piston overall length (L1) to the piston diameter (D) is in the range of 0.8 to 1.0, and when the length of the area where the piston seals the compression chamber by its reciprocating motion is defined as a sealing length (L2), the ratio L2/L1 of the sealing length (L2) to the piston overall length (L1) is in the range of 0.9 to 1.0.

In any of the above-mentioned configurations of the hermetic refrigerant compressor, the compression element may comprise a crankshaft having a main shaft and an eccentric shaft as a shaft portion, and a main bearing supporting the main shaft and an eccentric bearing supporting the eccentric shaft as a bearing portion supporting the shaft portion, the sliding surface of the main shaft with the main bearing being divided into a plurality of surfaces, and when the sum of the axial lengths of the plurality of sliding surfaces is the total sliding length Tt, the ratio Tt/K of the total sliding length Tt to the outer diameter K of the main shaft may be 1.26 or less.

According to the above configuration, by applying the configuration for setting the ratio Tt/K and the configuration for using a sulfur-based sliding property modifier, the main shaft sliding part consisting of the main shaft and main bearing can be well lubricated and wear of the main shaft sliding part can be well suppressed even in a state in which the sliding area is reduced by using a low-viscosity lubricant. As a result, the reliability of the refrigerant compressor can be further improved. Moreover, even when the operating frequency is a low rotation speed of 16 r/s or more and 35 r/s or less, good wear resistance can be achieved even if the supply amount of refrigeration oil is reduced. Therefore, an increase in sliding loss in the main shaft sliding part can also be suppressed, and a good coefficient of performance (COP) can be achieved.

In any of the above-mentioned configurations of the hermetic refrigerant compressor, the compression element includes a crankshaft having a main shaft and an eccentric shaft as a shaft portion, and a main bearing supporting the main shaft and an eccentric bearing supporting the eccentric shaft as a bearing portion supporting the shaft portion, and the sliding surface of the main shaft with the main bearing is a single surface or is divided into multiple surfaces, and when the sliding surface is a single surface, the axial length of the sliding surface is taken as a single sliding length T, or when the sliding surface is divided into multiple surfaces, the axial length of the sliding surface with the smallest axial length is taken as a single sliding length T, and the ratio T/K of the single sliding length T to the outer diameter K of the main shaft is 0.51 or less, and further, the refrigeration oil may be configured to contain sulfur or a compound having sulfur as a sliding property modifier.

According to the above configuration, by applying the configuration for setting the ratio T/K and the configuration for using a sulfur-based sliding property modifier, the main shaft sliding part consisting of the main shaft and main bearing can be well lubricated and wear of the main shaft sliding part can be well suppressed, even in a state in which the sliding area is reduced by using a low-viscosity lubricant. As a result, the reliability of the refrigerant compressor can be further improved. Moreover, even when the operating frequency is a low rotation speed of 16 r/s or more and 35 r/s or less, good wear resistance can be achieved even if the supply amount of refrigeration oil is reduced. Therefore, an increase in sliding loss in the main shaft sliding part can also be suppressed, and a good coefficient of performance (COP) can be achieved.

In any of the above configurations of the hermetic refrigerant compressor, the compression element may further include a crankshaft having a main shaft and an eccentric shaft, a main bearing supporting the main shaft, and a thrust bearing provided on the thrust surface of the main bearing, and the end of the sliding surface of the main bearing on the compression chamber side is a first end and the opposite end is a second end, the distance between the axis of the compression chamber and the second end of the sliding surface of the main bearing is P, and the distance between the axis of the compression chamber and the first end of the sliding surface of the main bearing is Q, and when the distance P is within the range of 38 mm to 51 mm, the distance Q is 16 mm or less.

With the above configuration, the load on the main shaft can be reduced simply by using low-viscosity oil as the refrigeration oil, but the load on the main shaft can be reduced even further by setting the distance Q of the refrigerant compressor to 16 mm or less. This makes it possible to achieve high efficiency and good reliability of the refrigerant compressor not only in the sliding part between the piston and cylinder but also in the sliding part of the main shaft. This makes it possible to further improve the coefficient of performance (COP) of the refrigerant compressor.

The present disclosure also includes a method for operating a hermetic refrigerant compressor. That is, the method for operating a hermetic refrigerant compressor according to the present disclosure may be configured such that, in a hermetic refrigerant compressor including a hermetic container, refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 mm ² /s, a cylinder block that is housed in the hermetic container and forms a compression chamber, and a piston that is inserted and reciprocatable within the compression chamber, when the operating frequency is 16 r/s or more and 35 r/s or less, the average reciprocating speed of the piston exceeds 0.31 m/s.

The present disclosure also includes a refrigeration/refrigeration device using a hermetic refrigerant compressor of the above configuration, or a hermetic refrigerant compressor that executes the operating method of the above configuration. In other words, the refrigeration/refrigeration device according to the present disclosure may be configured to include a hermetic refrigerant compressor of the above configuration (or a hermetic refrigerant compressor that executes the operating method of the above configuration), a radiator, a pressure reducing device, and a heat absorber, and to include a refrigerant circuit in which these are connected in a ring shape by piping.

Below, representative embodiments of the present disclosure will be described in detail with reference to the drawings. However, some detailed descriptions of the following embodiments may be omitted. For example, detailed descriptions of matters that are already well known, or duplicate descriptions of substantially identical configurations may be omitted. This is to avoid making the following description unnecessarily redundant and to make it easier for those skilled in the art to understand.

The accompanying drawings and the following description are provided to enable those skilled in the art to fully understand the present disclosure, and are not intended to limit the subject matter described in the claims.

(Embodiment 1)
[Configuration of hermetic refrigerant compressor]
First, a representative configuration example of a hermetic refrigerant compressor according to the present disclosure will be specifically described with reference to Fig. 1. Fig. 1 is a schematic cross-sectional view showing an example of the configuration of a hermetic refrigerant compressor 100 (hereinafter, may be abbreviated as refrigerant compressor 100) according to a first embodiment of the present disclosure.

As shown in FIG. 1, the refrigerant compressor 100 has a sealed container 102 filled with, for example, R600a as a refrigerant gas 181, and mineral oil is stored at the bottom as refrigeration oil 180. A compressor body 108 is housed in the sealed container 102, and this compressor body 108 is elastically supported by a suspension spring 190. The compressor body 108 also includes an electric element 104 and a compression element 106.

The electric element 104 is composed of at least a stator 150 and a rotor 152. The compression element 106 is a reciprocating type driven by the electric element 104, and includes a crankshaft 120, a cylinder block 130, a piston 140, a connecting means 142, etc. The crankshaft 120 is composed of at least a main shaft 124 to which the rotor 152 is shrink-fitted, and an eccentric shaft 122 formed eccentrically with respect to the main shaft 124. In this embodiment, the crankshaft 120 is composed of, for example, an iron-based material.

Furthermore, a flange portion 128 is provided between the main shaft 124 and the eccentric shaft 122. As shown by the dashed line in FIG. 1, the rotation axis of the crankshaft 120 corresponds to the axis of the main shaft 124. The main shaft 124 and the eccentric shaft 122 are fixed via the flange portion 128 so that their axes are misaligned. Therefore, the axis of the eccentric shaft 122 is eccentric with respect to the axis of the main shaft 124 (the axis of rotation of the crankshaft 120).

The eccentric shaft 122 of the crankshaft 120 is located above the refrigerant compressor 100, and the main shaft 124 is located below the refrigerant compressor 100. Therefore, this up-down positional relationship (direction) is also used when explaining the position of the crankshaft 120. For example, the upper end of the eccentric shaft 122 faces the inner upper surface of the sealed container 102, and the lower end of the eccentric shaft 122 is connected to the main shaft 124.

The upper end of the main shaft 124 is connected to the eccentric shaft 122, and the lower end of the main shaft 124 faces the inner bottom surface of the sealed container 102, and is immersed in refrigeration oil 180. In addition, the crankshaft 120 is provided with an oil supply mechanism 125, which supplies refrigeration oil 180 from the lower end of the main shaft 124 immersed in the refrigeration oil 180 to the upper end of the eccentric shaft 122. As described below, the refrigeration oil 180 lubricates each sliding part of the refrigerant compressor 100 and also provides a seal between the compression chamber 133 and the piston 140.

The outer peripheral surface of the main shaft 124 of the crankshaft 120 includes sliding

surfaces

126a, 126b and a non-sliding outer peripheral surface 127. For ease of explanation, the upper sliding surface 126a of the main shaft 124 is referred to as the first sliding surface 126a, and the lower sliding surface 126b of the main shaft 124 is referred to as the second sliding surface 126b. The non-sliding outer peripheral surface 127 is located between the first sliding surface 126a and the second sliding surface 126b.

In this disclosure, a "sliding surface" refers to the outer or inner peripheral surface of a plurality of sliding members constituting the sliding portion, which is in sliding contact with the inner or outer peripheral surface of the other. A "non-sliding outer peripheral surface (non-sliding surface)" is a surface that, unlike a sliding surface, does not contact the inner or outer peripheral surface of the other. In this embodiment, non-sliding outer peripheral surface 127 is configured such that the outer diameter of main shaft 124 is smaller than sliding

surfaces

126a, 126b (the outer diameter is narrowed, it is recessed from sliding

surfaces

126a, 126b, or it is formed by hollowing out).

The cylinder block 130 includes a cylinder 132 and a main bearing 134. The cylinder 132 defines a compression chamber 133 therein. The main bearing 134 supports the main shaft 124 so that it can rotate freely. In this embodiment, the cylinder 132 and the main bearing 134 are integrally formed as one cylinder block 130, for example, from cast iron.

In this embodiment, as shown in FIG. 1, if the extension direction (up-down direction) of the crankshaft 120 is defined as the "vertical direction", the cylinder block 130 has a main body that extends in the "horizontal direction" (direction perpendicular to the vertical direction) inside the refrigerant compressor 100. The main bearing 134 is formed in a tubular (cylindrical) shape that extends in the "vertical direction" (up-down direction) relative to the main body of the cylinder block 130. The inner peripheral surface of the main bearing 134 is in slidable contact with the outer peripheral surface of the main shaft 124, i.e., the sliding

surfaces

126a, 126b. Therefore, the inner peripheral surface of the main bearing 134 is a sliding surface.

The non-sliding outer peripheral surface 127 of the main shaft 124 is located between the upper and lower ends of the main bearing 134. Therefore, when the main shaft 124 is supported by the main bearing 134, the upper end of the main bearing 134 contacts the first sliding surface 126a of the main shaft 124, and the lower end of the main bearing 134 contacts the second sliding surface 126b. At this time, the non-sliding outer peripheral surface 127, which has a smaller outer diameter than the sliding

surfaces

126a and 126b, does not contact the inner peripheral surface (sliding surface) of the main bearing 134, and is not exposed from the upper and lower ends of the main bearing 134.

In this embodiment, the main bearing 134 has a thrust surface 136 and a tubular extension 137. The thrust surface 136 of the main bearing 134 is a flat surface that extends in a direction perpendicular to the axis, i.e., the extension direction (up-down direction) of the main shaft 124 (vertical direction, horizontal direction).

The tubular extension 137 of the main bearing 134 is tubular (cylindrical) and extends further upward than the thrust surface 136; in other words, it is a portion that extends upward from the tubular main bearing 134 body. Therefore, the tubular extension 137, together with the main bearing 134 body, has an inner circumferential surface (sliding surface) that faces the outer circumferential surface (sliding surface) of the main shaft 124. A thrust ball bearing 210 is provided on the thrust surface 136 of the main bearing 134.

Cylinder 132 is provided in the main body of cylinder block 130, and the inside of cylinder 132 forms compression chamber 133. Compression chamber 133 is a cylindrical (columnar) bore that extends "horizontally" inside refrigerant compressor 100. Piston 140 is inserted into this compression chamber 133 so that it can reciprocate. Therefore, compression chamber 133 is closed by the insertion of piston 140. Furthermore, the direction in which piston 140 reciprocates is the "horizontal direction".

The connecting means 142 is made of, for example, an aluminum casting, supports the eccentric shaft 122, and is connected to the piston 140. Therefore, the eccentric shaft 122 and the piston 140 are connected by the connecting means 142.

The electric element 104 includes a rotor 152 and a stator 150 arranged coaxially with the rotor 152 so as to surround the rotor 152. The stator 150 is arranged on the outer diameter side of the rotor 152 so as to maintain a substantially constant gap with the rotor 152, and is fixed to the legs of the cylinder block 130. The rotor 152 is also fixed to the main shaft 124 of the crankshaft 120.

Therefore, in the refrigerant compressor 100, when the rotor 152 is rotated by the electric element 104, the crankshaft 120 rotates. As described above, in the crankshaft 120, the axis of the eccentric shaft 122 is offset from the axis of the main shaft 124, and the eccentric shaft 122 is connected to the piston 140 by a connecting means 142. The piston 140 is inserted in the compression chamber 133 in the cylinder 132 so that it can reciprocate. Therefore, when the crankshaft 120 rotates, the rotation of the eccentric shaft 122 causes the piston 140 to reciprocate within the compression chamber 133.

Furthermore, as described above, as the crankshaft 120 rotates, refrigeration oil 180 is supplied from the oil supply mechanism 125 to each sliding part. As a result, each sliding part is lubricated by the refrigeration oil 180.

In this embodiment, the sliding parts include the main shaft 124 and main bearing 134 of the crankshaft 120, the piston 140 and the compression chamber 133 (cylinder 132), the connecting means 142 and the connecting part of the piston 140, and the connecting part of the eccentric shaft 122 of the crankshaft 120 and the connecting means 142. Each of the members that make up these sliding parts is a sliding member.

Furthermore, for the sake of convenience, the sliding part formed by the piston 140 and the compression chamber 133 (cylinder 132) among these sliding parts is referred to as the "cylinder sliding part." Furthermore, for the sake of convenience, the sliding part formed by the main shaft 124 and main bearing 134 of the crankshaft 120 is referred to as the "main shaft sliding part."

The specific configuration of the refrigerant compressor 100 according to the present disclosure is not limited to the configuration shown in FIG. 1. The refrigerant compressor 100 according to the present disclosure may be configured to include an electric element 104 and a compression element 106, and the compression element 106 may include a cylinder block 130 having a compression chamber 133, and a piston 140 inserted in the compression chamber 133 so as to be capable of reciprocating motion.

For example, in the refrigerant compressor 100 shown in FIG. 1, the electric element 104 is located on the lower side and the compression element 106 is located on the upper side within the sealed container 102. However, the electric element 104 may be located on the upper side and the compression element 106 may be located on the lower side.

Alternatively, in the refrigerant compressor 100 shown in FIG. 1, the electric element 104 is an inner rotor type, and the rotor 152 is rotatably arranged coaxially with the stator 150 on the inner circumference of the stator 150. However, the configuration of the electric element 104 is not limited to this, and it may be an outer rotor type, i.e., the rotor 152 is rotatably arranged coaxially with the stator 150 on the outer circumference of the stator 150.

Alternatively, in the refrigerant compressor 100 shown in FIG. 1, the main shaft 124 of the crankshaft 120 has a first sliding surface 126a, a non-sliding outer circumferential surface 127, and a second sliding surface 126b. However, the configuration of the main shaft 124 is not limited to this, and as in the second embodiment described below, the sliding surface 126 of the main shaft 124 may constitute the entire outer circumferential surface of the main shaft 124, or the main shaft 124 may have three or more sliding surfaces 126.

The specific configuration of the refrigeration oil 180 used in the present disclosure is not particularly limited. In the present disclosure, it is sufficient that an oil (low-viscosity oil) having a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 mm ² /s (1.0 mm ² /s or more and 2.5 mm ² /s or less) is used as the refrigeration oil 180. The specific configuration of the low-viscosity oil is not particularly limited, and in the first embodiment, for example, a low-viscosity mineral oil is used, but as described in the fourth embodiment described later, another oily substance may be used instead of the mineral oil, another oily substance may be used in combination with the mineral oil, or various additives and the like may be added.

In addition, when the kinetic viscosity of the refrigeration oil 180 (low viscosity oil) in this disclosure at 40° C. exceeds 2.5 mm ² /s, the viscous resistance becomes too large from the viewpoint of realizing a good coefficient of performance (COP), and the input power to the refrigerant compressor 100 becomes large when the viscous resistance becomes large, making it impossible to realize a better coefficient of performance (COP).

On the other hand, if the kinetic viscosity of the refrigeration oil 180 at 40°C is less than 1.0 ^mm2 /s, the oil film formed on each sliding part inside the refrigerant compressor 100 becomes too thin. If the oil film becomes thin at the sliding parts, the possibility of breakage increases, and the sliding parts cannot obtain a satisfactory lubricating effect. This increases metal contact between the sliding surfaces at the sliding parts, and there is a possibility that the reliability of the sliding parts will decrease.

Here, the kinematic viscosity of the refrigeration oil 180 according to the present disclosure at 40° C. may be within the above-mentioned range, but the upper and lower limits may be changed as appropriate within the above-mentioned range depending on various conditions. For example, the upper limit of the kinematic viscosity of the refrigeration oil 180 at 40° C. may be 2.4 mm ² /s. Moreover, the lower limit of the kinematic viscosity of the refrigeration oil 180° C. at 40° C. may be 1.5 mm ² /s.

In addition, these upper and lower limit values may be values including (less than or equal to) the numerical value, or may be values not including (less than or greater than) the numerical value. For example, the upper limit value may be less than 2.5 mm ² /s, less than 2.4 mm ² /s, or less than 2.4 mm ² /s. The same applies to the lower limit value. Depending on various conditions, by setting the kinetic viscosity of the refrigeration oil 180 at 40°C to less than or equal to 2.4 mm ² /s, or to more than or equal to 1.5 mm ² /s, it becomes easier to realize a more suitable viscous resistance or a more suitable oil film thickness from the viewpoint of realizing a better coefficient of performance (COP).

[Refrigerant compressor operation method and piston configuration]
Next, the operation method of the hermetic refrigerant compressor according to the present disclosure and the piston configuration in the hermetic refrigerant compressor will be described with reference to the above-mentioned refrigerant compressor 100 shown in Fig. 1. Note that the "piston configuration" here includes not only the specific configuration of the piston 140 itself but also the conditions set when the piston 140 reciprocates within the compression chamber 133.

In the refrigerant compressor 100, first, power is supplied from a commercial power source to the electric element 104, causing the rotor 152 of the electric element 104 to rotate. The rotation of the rotor 152 causes the crankshaft 120 to rotate as described above, and the eccentric motion of the eccentric shaft 122 relative to the main shaft 124 is transmitted to the piston 140 via the connecting means 142. As a result, the eccentric motion of the eccentric shaft 122 is converted into the reciprocating motion of the piston 140, and the piston 140 is driven to reciprocate inside the cylinder 132, i.e., within the compression chamber 133. The refrigerant gas 181 introduced into the sealed container 102 is sucked into the compression chamber 133 by the reciprocating motion of the piston 140 and compressed.

Here, in the present disclosure, it is preferable that the refrigerant compressor 100 is inverter-driven at multiple operating frequencies. In other words, the refrigerant compressor 100 according to the present disclosure may include at least an inverter circuit that controls the operating frequency as a controller that controls the operation of the refrigerant compressor 100. The specific configuration of the inverter circuit is not particularly limited as long as it can rotate the electric element 104 at multiple operating speeds. The inverter circuit may be formed into a chip, or may be a microprocessor or the like that operates according to a program that executes rotational drive at multiple operating speeds.

In the present disclosure, the operating frequency of the refrigerant compressor 100 is not particularly limited. In general, lowering the operating frequency of the refrigerant compressor 100 can reduce its power consumption. However, in a refrigeration cycle (freezing/refrigeration device) equipped with the refrigerant compressor 100, the operating frequency may also increase depending on the peak of the refrigeration capacity. Therefore, the operating frequency of the refrigerant compressor 100 is not always within a low range.

Here, in the present disclosure, the refrigeration oil 180 stored in the sealed container 102 of the refrigerant compressor 100 is a low-viscosity oil (low-viscosity lubricating oil) having a kinetic viscosity at 40°C in the range of 1.0 mm ^{2 /} s to 2.5 mm ² /s.

When low-viscosity oil is used as the refrigeration oil 180, it is possible to effectively reduce sliding losses, but the lubricating effect at the sliding parts tends to decrease. Therefore, when the refrigeration oil 180 is a low-viscosity oil, a method of avoiding or suppressing the decrease in lubricating effect is generally chosen that involves, for example, applying a surface treatment to the sliding surfaces that make up the sliding parts.

Furthermore, as mentioned above, according to the studies of the present inventors, it has become clear that if the kinetic viscosity of refrigeration oil 180 at 40°C is 2.5 ^mm2 /s or less, the amount of leakage of refrigerant gas 181 from between piston 140 and cylinder 132 (between the outer peripheral surface of piston 140 and the inner peripheral surface of cylinder 132 (the inner peripheral surface of compression chamber 133)) becomes too large (see also the examples described later). As a method for suppressing the amount of leakage of refrigerant gas 181, for example, a measure such as reducing the clearance between the surface of piston 140 and the inner surface of compression chamber 133 is generally selected.

In contrast, in the present disclosure, when a low-viscosity oil having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to 2.5 ^mm2 /s is used as the refrigeration oil 180 stored in the sealed container 102, and when the operating frequency is controlled by the inverter circuit (or controller) to be 16 r/s or more and 35 r/s or less, the refrigerant compressor 100 is operated so that the average reciprocating speed of the piston 140 (average piston speed) exceeds 0.31 m/s.

By using such an operating method, even if a low-viscosity oil with a kinetic viscosity of 2.5 ^mm2 /s or less at 40°C is used as refrigerating machine oil 180, the oil film of refrigerating machine oil 180 can exhibit good viscous force (viscosity resistance) between piston 140 and cylinder 132 in the "cylinder sliding portion," that is, the sliding portion formed by piston 140 and compression chamber 133 in cylinder 132. This makes it possible to achieve good sliding performance in the cylinder sliding portion and effectively suppress leakage of refrigerant gas 181.

As a result, even when the refrigerant compressor 100 is operated within a low operating frequency range, i.e., when the operating frequency is controlled within a range of 16 r/s to 35 r/s, it is possible to suppress or avoid an increase in the power consumption of the refrigerant compressor 100, and it is possible to achieve a good coefficient of performance (COP) in the refrigerant compressor 100.

On the other hand, if the oil film of the refrigeration oil 180 cannot exert a good viscous force in the cylinder sliding part, it becomes difficult to maintain a good oil film between the piston 140 and the cylinder 132 against the pressure of the refrigerant gas 181 in the compression chamber 133. This makes it easier for the oil film to break between the piston 140 and the cylinder 132, resulting in a significant amount of refrigerant gas 181 leaking.

The viscous force of the oil film of the refrigeration oil 180 in the cylinder sliding part will be specifically described with reference to FIG. 2. FIG. 2 is a schematic side view (schematic partial cross-sectional view) showing an enlarged view of a main part of the configuration of the cylinder sliding part provided in the refrigerant compressor 100 shown in FIG. 1.

2 shows a schematic diagram of a part of a cylinder 132 provided in a cylinder block 130, a part of a compression chamber 133 formed in the cylinder 132, and a part of a piston 140 slidably inserted in the compression chamber 133. The compression chamber 133 is filled with a refrigerant gas 181, and an oil film of refrigerating machine oil 180 is formed between the outer peripheral surface of the piston 140 and the inner peripheral surface of the cylinder 132 (the inner peripheral surface of the compression chamber 133).

When the viscous force (F1) of the oil film of the refrigerating machine oil 180 present between the piston 140 and the cylinder 132 is taken as the viscous force, the viscous force (F1) can be expressed by the following formula (1) using the viscosity (η) of the refrigerating machine oil 180, the seal length (L2) of the piston 140, the clearance (σ) between the piston 140 and the cylinder 132, and the average piston speed (V).
F1=(η×L2×V)/σ (1)

The seal length (L2) is the length of the area where the piston 140 seals the compression chamber 133 by its reciprocating motion. The viscous force of the oil film (F1) is a force that acts from the outside to the inside of the compression chamber 133, as shown diagrammatically by the block arrow in FIG. 2. The clearance (σ) between the piston 140 and the cylinder 132 is also shown diagrammatically in FIG. 2, and ideally, the oil film of the refrigeration oil 180 fills the clearance (σ) without breaking.

If the viscosity force of the oil film (F1) becomes smaller, it becomes difficult for the refrigeration oil 180 to maintain the state of an oil film between the piston 140 and the cylinder 132 against the pressure of the refrigerant gas 181 filling the compression chamber 133. This makes it easier for the oil film to break between the piston 140 and the cylinder 132, and as a result, it becomes easier for a significant amount of refrigerant gas 181 to leak.

Therefore, in order to increase the viscous force of the oil film (F1), it is possible to increase (speed up) the average piston speed (V), increase the seal length (L2) of the piston 140, and reduce the clearance (σ) between the piston 140 and the cylinder 132.

In this disclosure, when the refrigeration oil 180 is a low viscosity oil, the piston average speed (V) is set to a large value in order to exert a good oil film viscous force between the piston 140 and the cylinder 132. Because the piston 140 reciprocates in the cylinder sliding portion, the speed at which the piston 140 actually moves within the compression chamber 133 is not constant. Therefore, in this disclosure, the piston average speed is used. The piston average speed (V) can be defined as the product (V = S x Fr) of the stroke amount (S) of the piston 140 and the operating frequency (Fr).

In the present disclosure, the operating frequency (operating rotation speed) of the refrigerant compressor 100 is not particularly limited, but typically, the lower limit of the operating frequency can be 13 r/s (rps), which may be 16 r/s. On the other hand, the upper limit of the operating frequency can be 80 r/s, which may be 75 r/s.

Therefore, a typical operating frequency range can be in the range of 13 to 80 r/s, and may be in the range of 16 to 75 r/s. Of course, it may be in the range of 16 to 80 r/s, or in the range of 13 r/s to 75 r/s. Note that although the input power to the refrigerant compressor 100 will be large, the upper limit of the operating frequency may exceed 80 r/s.

Furthermore, from the viewpoint of reducing the input power to the refrigerant compressor 100, the operating frequency may be set lower than 13 r/s. However, if the operating frequency is set too low, there is a possibility that sufficient reliability will not be obtained in terms of the coefficient of performance (COP) of the refrigerant compressor 100 and the wear of the cylinder sliding parts. Therefore, a preferable lower limit of the operating frequency can be set to 16 r/s.

In this disclosure, within this range of operating frequencies, the range of 16 r/s or more and 35 r/s or less is specifically defined as the "low-speed operating frequency." In the refrigerant compressor 100 according to the present disclosure, when within this range of low-speed operating frequencies, the average piston speed is increased to exceed 0.31 m/s. Therefore, the lower limit of the average piston speed in this disclosure only needs to exceed 0.31 m/s.

If the average piston speed is 0.31 m/s or less, the oil film of the low-viscosity oil (refrigeration oil 180) cannot exert sufficient viscous force between the piston 140 and the cylinder 132. As a result, when the refrigerant compressor 100 is operating within the low-speed operating frequency range, the refrigerant gas 181 is likely to leak from between the piston 140 and the cylinder 132.

As mentioned above, the lower limit of the average piston speed should be greater than 0.31 m/s, but depending on various conditions, it may be greater than 0.32 m/s or greater than 0.34 m/s. If the lower limit of the average piston speed is greater than 0.32 m/s, the oil film of the refrigeration oil 180 between the piston 140 and the cylinder 132 will be able to exert even better viscous forces, depending on various conditions. This makes it possible to better suppress leakage of the refrigerant gas 181 from between the piston 140 and the cylinder 132.

In the present disclosure, a configuration can be adopted in which the ratio S/D of the reciprocating stroke amount (S) of piston 140 to the piston diameter (D) is set within the range of 0.78 to 1.00 based on the above formula (1) (0.78≦S/D≦1.00). The reciprocating stroke amount (S) of piston 140 is determined by twice the eccentric radius of eccentric shaft 122. This ratio S/D corresponds to the condition set when piston 140 reciprocates within compression chamber 133, and can therefore be said to be the piston configuration described above.

In this way, by setting the ratio S/D of stroke amount (S) to piston diameter (D) within the above range, the stroke amount of piston 140 can be relatively longer (larger), and therefore the average piston velocity (V) can be increased. As described above, if the average piston velocity (V) is increased, the viscous force (F1) of the oil film can be increased, making it easier to form an oil film between piston 140 and cylinder 132. This makes it possible to suppress leakage of refrigerant gas 181.

Furthermore, when the ratio S/D is within the above range, the fact that the stroke amount (S) is large means that the piston diameter (D) is relatively small. By reducing the piston diameter (D), the total area of the clearance (σ) between the piston 140 and the cylinder 132 can be reduced. Reducing the total area of the clearance means that the "opening area" from which the refrigerant gas 181 may leak is narrowed. Therefore, it is possible to suppress leakage of the refrigerant gas 181 from between the piston 140 and the cylinder 132.

Furthermore, if the piston diameter (D) is reduced, the compression load of the refrigerant gas 181 by the piston 140 can be reduced. In other words, when the piston 140 reciprocates in the compression chamber 133, the load applied to the tip surface of the piston 140 from the refrigerant gas 181 in the compression chamber 133 becomes relatively smaller. This makes it possible to reduce the input power for reciprocating the piston 140. Therefore, the coefficient of performance (COP) of the refrigerant compressor 100 can be improved.

Until now, in the field of refrigerant compressors 100, as disclosed in "Sealed Refrigeration Machines" by Mutsuyoshi Kawahira, published by the Japan Refrigeration Association in July 1981, it has been considered desirable for the ratio S/D of the stroke amount (S) to the piston diameter (D) to be in the range of 0.4 to 0.8. In contrast, in this disclosure, the ratio S/D is set in the range of 0.78 to 1.00. Therefore, the ratio S/D in this disclosure is set to be substantially larger than the conventional range. In particular, in this disclosure, the lower limit of the ratio S/D may be 0.81 or more (0.81≦S/D) or 0.84 or more (0.84≦S/D).

If the ratio S/D is less than 0.78, when a low-viscosity oil (with a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to 2.5 ^mm2 /s) is used as the refrigeration oil 180, it may become impossible to make the stroke amount (S) of the piston 140 relatively large, and it may become impossible to make the piston diameter (D) relatively small. If the lower limit of the ratio S/D is 0.81 or more, the effect of being able to relatively increase the stroke amount (S) and relatively reduce the piston diameter (D) can be more reliably achieved. If the lower limit of the ratio S/D is 0.84 or more, the above-mentioned effect can be more reliably achieved.

On the other hand, if the ratio S/D exceeds 1.00, the stroke amount (S) of the piston 140 becomes relatively too large, and the sliding loss between the piston 140 and the cylinder 132 becomes relatively large. As a result, the effect of realizing a good coefficient of performance (COP) in the refrigerant compressor 100 cannot be obtained.

Note that, although there are no particular limitations on the specific stroke amount (S) of the piston 140, in this embodiment, for example, the lower limit of the stroke amount (S) can be 19.5 mm or more. The lower limit of the stroke amount (S) may be 20 mm or more. On the other hand, there are no particular limitations on the upper limit of the stroke amount (S), but if the stroke amount (S) becomes excessively large, there is a risk that the sliding loss between the piston 140 and the cylinder 132 will become relatively large. From this perspective, the upper limit of the stroke amount (S) can be 30 mm or less.

Furthermore, the clearance (σ) between the piston 140 and the cylinder 132 is not particularly limited to a specific distance, but in this embodiment, for example, the lower limit of the clearance (σ) can be 3 μm, and the upper limit of the clearance (σ) can be 10 μm. If the clearance (σ) exceeds 10 μm, the viscous force (F1) becomes small, especially when a low-viscosity oil is used as the refrigeration oil 180 (see formula (1) above).

On the other hand, if the clearance (σ) is less than 3 μm, there is a risk that the reciprocating piston 140 will easily come into contact with the inner surface of the cylinder 132 (compression chamber 133), especially when a low-viscosity oil is used as the refrigeration oil 180.

In the present disclosure, based on the above formula (1), a configuration can be adopted in which the ratio L1/D of the piston overall length (L1) to the piston diameter (D) of the piston 140 is within the range of 0.8 to 1.0 (0.8≦L1/D≦1.0), and the ratio L2/L1 of the seal length (L2) to the piston overall length (L1) is within the range of 0.9 to 1.0 (0.9≦L2/L1≦1.0).

Among these ratios, the ratio L1/D corresponds to the specific configuration (condition) of the piston 140, and the ratio L2/L1 corresponds to the condition set when the piston 140 reciprocates within the compression chamber 133. Therefore, all of these ratios can be said to be the piston configuration described above.

The relationship between the piston diameter (D), overall piston length (L1), and seal length (L2) in the piston 140 is defined by the ratios L1/D and L2/L1, as will be specifically explained with reference to Figures 3A and 3B. Figure 3A is a schematic side view showing a representative example of the piston 140 used in the refrigerant compressor 100 shown in Figure 1, and Figure 3B is a schematic side view showing a representative example of a conventional piston.

As shown in FIG. 3A, in the piston 140 used in the refrigerant compressor 100 according to this embodiment, the piston overall length (L1), i.e., the length along the direction in which the piston 140 reciprocates, is approximately equal to the piston diameter (D), i.e., the diameter of the piston 140. In addition, in the piston 140, the piston overall length (L1) is also approximately equal to the seal length (L2), i.e., the length of the area that seals the inside of the compression chamber 133 due to the reciprocating motion of the piston 140 described above.

In contrast, as shown in FIG. 3B, in the conventional piston 240, the overall piston length (L1) is larger than the piston diameter (D), and the overall piston length (L1) is larger than the seal length (L2).

As is clear from the above formula (1), by increasing the seal length (L2), the viscous force (F1) of the oil film of the refrigeration oil 180 can be increased. Furthermore, if the seal length (L2) is large, this means that the area sealed by the oil film in the cylinder sliding part becomes larger. Therefore, if the seal length (L2) is relatively large, this alone can effectively suppress leakage of the refrigerant gas 181.

In order to ensure a sufficient seal length (L2), the total piston length (L1) must also be sufficiently large. On the other hand, if the total piston length (L1) is too large, sliding loss at the cylinder sliding parts increases. In particular, when the refrigeration oil 180 is a low-viscosity oil (kinematic viscosity at 40°C of 2.5 ^mm2 /s or less), it is more difficult to form a good oil film at the cylinder sliding parts compared to oils with higher viscosity.

As a result of thorough research into the criteria for setting the appropriate piston overall length (L1), it became clear that the piston diameter (D) should be used as the criterion, particularly when the refrigeration oil 180 is a low-viscosity oil.

As with the piston 140 shown in Figure 3A, if the ratio L1/D is within the range of 0.8 to 1.0, the range of the overall piston length (L1) suitable for low-viscosity oil is defined in the cylinder sliding part. This makes it possible to realize an overall piston length (L1) that can ensure an excellent seal length (L2) while suppressing an increase in sliding loss.

If the ratio L1/D is less than 0.8, when low-viscosity oil is used as the refrigeration oil 180, the overall piston length (L1) becomes relatively short. As a result, a sufficient seal length (L2) cannot be ensured, and not only does the size of the area sealed by the oil film become insufficient, but the viscous force (F1) of the oil film based on the above formula (1) cannot be increased.

On the other hand, if the ratio L1/D exceeds 1.0, as in the conventional piston 240 shown in FIG. 3B, when a low-viscosity oil is used as the refrigeration oil 180, the overall piston length (L1) becomes too large. This may lead to an increase in sliding loss in the cylinder sliding parts. If the sliding loss increases, it becomes difficult to achieve a good coefficient of performance (COP).

In this disclosure, since low viscosity oil is used as the refrigeration oil 180, it is desirable for the seal length (L2) to be as large as possible. To achieve this, it is desirable to also increase the overall piston length (L1) as described above. However, as described above, if the overall piston length (L1) is too large, sliding loss in the cylinder sliding parts increases.

Therefore, in this disclosure, the ratio L2/L1 is set within the range of 0.9 to 1.0. This allows the seal length (L2) to be relatively large without excessively increasing the overall piston length (L1), as in the piston 140 shown in FIG. 3A, for example. As a result, as is clear from the above formula (1), it is possible to increase the viscous force (F1) of the oil film while suppressing an increase in sliding loss in the cylinder sliding parts.

Moreover, the fact that the seal length (L2) is relatively large compared to the overall piston length (L1) means that the area sealed by the oil film of the refrigeration oil 180 is also relatively large. This makes it possible to further suppress leakage of the refrigerant gas 181. Furthermore, if the seal length (L2) is large, it becomes possible to stabilize the posture of the piston 140 reciprocating within the compression chamber 133. As a result, it becomes possible to further suppress an increase in sliding loss.

In contrast, if the ratio L2/L1 is less than 0.9, as in the conventional piston 240 shown in FIG. 3B, when low-viscosity oil is used as the refrigeration oil 180, the seal length (L2) cannot be sufficiently secured. As a result, not only cannot the viscous force (F1) of the oil film be increased, but the area sealed by the oil film becomes relatively small. This may result in the risk of not being able to sufficiently suppress leakage of the refrigerant gas 181. In addition, the ratio L2/L1 cannot exceed 1.0 (because the seal length (L2) is equal to or less than the total piston length (L1)).

In the present disclosure, the outer peripheral surface of piston 140 may be a smooth surface without any intentional unevenness, but may have, for example, an annular oil supply groove formed therein. By forming an oil supply groove on the outer peripheral surface of piston 140, it becomes possible to supply a sufficient amount of refrigeration oil 180 to the seal area of piston 140, particularly when low viscosity oil is used as refrigeration oil 180.

The specific configuration of the annular oil groove is not particularly limited. For example, the number of oil grooves is not particularly limited, but one can be mentioned as a typical example. Of course, two or more oil grooves may be formed. The width of the oil groove is also not particularly limited, but can be within the range of 0.1 to 0.5 mm, for example.

If the width of the oil supply groove is less than 0.1 mm, it will be difficult to supply a sufficient amount of refrigeration oil 180 to the seal area, even if the refrigeration oil 180 is a low-viscosity oil. On the other hand, if the width of the oil supply groove is more than 0.5 mm, if the refrigeration oil 180 is a low-viscosity oil, the width of the oil supply groove will be too wide and the refrigeration oil 180 will flow out of the seal area, making it difficult to retain an appropriate amount of refrigeration oil 180 in the seal area.

Here, for ease of explanation, the configuration for setting the average reciprocating speed of piston 140 (piston average speed) to exceed 0.31 m/s when the rotation frequency is controlled to be 16 r/s or more and 35 r/s or less is abbreviated to "configuration for increasing piston average speed (V)", and the configuration for setting the ratio S/D of the reciprocating stroke amount (S) of piston 140 to piston diameter (D) within the range of 0.78 to 1.00 is abbreviated to "configuration for setting ratio S/D", and When the ratio L1/D of the piston overall length (L1) to the piston diameter (D) is set within the range of 0.8 to 1.0, and the ratio L2/L1 of the seal length (L2) to the piston overall length (L1) is set within the range of 0.9 to 1.0, the configuration for increasing the average piston speed (V), the configuration for setting the ratio S/D, and the configuration for setting the ratios L1/D and L2/L1 can each be applied independently to the refrigerant compressor 100.

In other words, in the refrigerant compressor 100 according to the present disclosure, by applying at least one of the configurations for increasing the average piston speed (V), the configuration for setting the ratio S/D, and the configuration for setting the ratios L1/D and L2/L1, it is possible to effectively suppress leakage of the refrigerant gas 181, even when a low-viscosity oil is used as the refrigeration oil 180, and to achieve an even better coefficient of performance (COP).

In this way, in the hermetic refrigerant compressor of this embodiment 1, when a low-viscosity oil having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to 2.5 ^mm2 /s is used as the refrigeration oil 180, if a controller for controlling the operating frequency of the refrigerant compressor 100 is provided, when the operating frequency is controlled by the controller to be 16 r/s or more and 35 r/s or less, the average speed at which the piston 140 reciprocates can be set to greater than 0.31 m/s.

Alternatively, in the hermetic refrigerant compressor according to the first embodiment, when the low viscosity oil is used as the refrigeration oil, the ratio S/D of the reciprocating stroke amount (S) of the piston 140 to the piston diameter (D) may be within the range of 0.78 to 1.00.

Alternatively, in the hermetic refrigerant compressor according to the first embodiment, when the low viscosity oil is used as the refrigeration oil 180, the length of the area where the piston 140 seals the inside of the compression chamber 133 by its reciprocating motion is defined as the sealing length (L2), and the ratio L1/D of the piston total length (L1) to the piston diameter (D) may be within the range of 0.8 to 1.0, and the ratio L2/L1 of the sealing length (L2) to the piston total length (L1) may be within the range of 0.9 to 1.0.

A hermetic refrigerant compressor with these configurations can further reduce leakage of refrigerant gas 181 from between piston 140 and compression chamber 133 when low viscosity oil is used as refrigeration oil 180. This makes it possible to further improve the coefficient of performance (COP) of the hermetic refrigerant compressor.

(Embodiment 2)
The hermetic refrigerant compressor according to the second embodiment has a basic configuration similar to that of the hermetic refrigerant compressor according to the first embodiment, but has a more characteristic configuration in a main shaft sliding portion (a sliding portion formed by the main shaft 124 of the crankshaft 120 and the main bearing 134). Since the basic configuration of the hermetic refrigerant compressor according to the second embodiment is similar to the configuration shown in Fig. 1 in the first embodiment, detailed description thereof will be omitted.

[Main shaft sliding part]
An example of a specific configuration of the main shaft sliding part in the present embodiment 2 will be specifically described with reference to Figures 4A to 4C. Figure 4A is a schematic diagram showing an example of the configuration when the sliding surface of the crankshaft 120 included in the refrigerant compressor 100 shown in Figure 1 is a single surface, and Figures 4B and 4C are schematic diagrams showing an example of the configuration when the sliding surface of the crankshaft 120 is divided into multiple surfaces.

In the hermetic refrigerant compressor shown in FIG. 1, the main shaft 124 of the crankshaft 120, which is the shaft portion, has a configuration having a first sliding surface 126a and a second sliding surface 126b, so that it can be said that the sliding surface of the main shaft 124 is divided into multiple surfaces. The configuration of the main shaft 124 shown in FIG. 1, i.e., the configuration in which the sliding surface is divided into two surfaces, corresponds to the schematic diagram shown in FIG. 4B. The shaft portion according to the present disclosure is not limited to this, and may be a single surface. For example, as shown in FIG. 4A, the outer circumferential surface of the main shaft 124 is not divided into multiple sliding surfaces, and may be configured to have only a single sliding surface 126.

The specific configuration for dividing the sliding surface into multiple parts is not particularly limited, but typically, a recess that is recessed (concave) toward the central axis side of the sliding surface is formed between the multiple sliding surfaces. This recess constitutes the non-sliding outer circumferential surface 127, as shown in Figures 1 and 4B. The specific shape of the recess is also not particularly limited, and for example, the depth may be any depth as long as it does not affect the rigidity, strength, etc. of the main shaft 124. Similarly, the width of the recess (i.e., the distance between the multiple sliding surfaces) is also not particularly limited, and can be set appropriately depending on the degree to which the width of the sliding surface (sliding area) is narrowed (lowered or reduced).

When the sliding surface is divided into a plurality of sliding surfaces, the number of sliding surfaces is not particularly limited. As shown in FIG. 1 and FIG. 4B, the sliding surface may be divided into a total of two surfaces, the first sliding surface 126a and the second sliding surface 126b, or as shown in FIG. 4C, the sliding surface may be divided into a total of three surfaces, the first sliding surface 126c, the second sliding surface 126d, and the third sliding surface 126e, or may be divided into four or more surfaces. In the configuration shown in FIG. 4C, the first non-sliding outer peripheral surface 127a, which is a recess similar to the non-sliding outer peripheral surface 127, is located between the first sliding surface 126c and the second sliding surface 126d, and the second non-sliding outer peripheral surface 127b is located between the second sliding surface 126d and the third sliding surface 126e.

In this second embodiment, the ratio of the axial length of the sliding surface to the outer diameter (diameter) of the portion that forms the sliding surface in the main shaft sliding portion is set to a predetermined value or less, thereby making it possible to reduce the sliding area without substantially affecting the wear resistance.

Specifically, when the sliding surface is a single surface (e.g., see FIG. 4A), the axial length of the sliding surface is the single sliding length T, and when the sliding surface is divided into multiple surfaces (e.g., FIG. 4B or FIG. 4C), the axial length of the sliding surface with the shortest axial length is the single sliding length T. Then, when the outer diameter (diameter) of the portion of the shaft that becomes the sliding surface is the outer diameter K, the shaft is designed so that the ratio T/K of the single sliding length T to the outer diameter K of the shaft is 0.51 or less.

In FIG. 4A, for the convenience of explaining the outer diameter K and the single sliding length T, the length T of the single sliding surface 126 (single sliding length T) is illustrated large relative to the outer diameter K, and if FIG. 4A is followed, the ratio T/K exceeds 0.51. However, in reality, the ratio T/K can be set to 0.51 or less (T/K≦0.51) by forming a recess (non-sliding outer peripheral surface) on the upper part (eccentric shaft 122 side) or lower part (refrigerant oil 180 side) of the main shaft 124 when viewed from the single sliding surface 126, for example.

In FIG. 4B, the sliding surface is divided into a first sliding surface 126a and a second sliding surface 126b. In the example shown in FIG. 4B, the axial length Ta of the upper first sliding surface 126a is smaller than the axial length Tb of the lower second sliding surface 126b (Ta<Tb). In this case, the first sliding surface 126a is the "shortest sliding surface," and its length Ta corresponds to the single sliding length T (T=Ta). In this example, it is sufficient that Ta/K is 0.51 or less for the first sliding surface 126a.

In FIG. 4B, as in FIG. 4A, the outer diameter K and the length Ta of the first sliding surface 126a are illustrated with the length Ta enlarged relative to the outer diameter K for ease of explanation. In this case as well, the ratio T/K can be set to 0.51 or less by increasing the axial length of the non-sliding outer peripheral surface 127 or by providing a non-sliding outer peripheral surface (recessed portion) (not shown) above the first sliding surface 126a.

In FIG. 4C, the sliding surface is divided into a first sliding surface 126c, a second sliding surface 126d, and a third sliding surface 126e. In the example shown in FIG. 4C, the length Td of the second sliding surface 126d in the center is smaller than the axial length Tc of the upper first sliding surface 126c, and the length Tc is smaller than the length Te of the lower third sliding surface 126e (Td<Tc<Te). In this case, the second sliding surface 126d is the "shortest sliding surface," and its length Td corresponds to the single sliding length T (T=Te). In this example, it is sufficient that Te/K is 0.51 or less for the second sliding surface 126d.

In this disclosure, the lower limit of the ratio T/K is not particularly limited, but a preferred example of the lower limit is 0.15 or more. Therefore, a preferred range of the ratio T/K in this disclosure is within the range of 0.15 to 0.51. A more preferred lower limit of the ratio T/K is 0.30, and an even more preferred lower limit is 0.42.

If the ratio T/K exceeds 0.51, when a low-viscosity oil (with a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 ^{mm 2} /s) is used as the refrigeration oil 180, sufficient wear resistance cannot be obtained even if a sulfur-based sliding property modifier described below is added to the refrigeration oil 180. On the other hand, if the ratio T/K is less than 0.15, depending on the conditions of the shaft portion, the sliding surface may become too narrow. In general, if the ratio T/K is 0.15 or more, the sliding area is not excessively reduced, so that even if a low-viscosity oil is used as the refrigeration oil 180, the wear resistance of the main shaft sliding portion can be suitably achieved by the sulfur-based sliding property modifier.

Alternatively, in this second embodiment, when the sliding surface of the main shaft sliding portion is divided into multiple surfaces, an axial length other than the single sliding length T described above may be specified, and the ratio of the axial length to the outer diameter (diameter) of the sliding surface may be set to a predetermined value or less. This allows the sliding area to be reduced without substantially affecting the wear resistance.

Specifically, in this second embodiment, when the sliding surface is divided into multiple surfaces, the shaft portion may be designed so that when the total axial length Tt is the sum of the axial lengths of the multiple sliding surfaces, the ratio Tt/K of the total sliding length Tt to the outer diameter K is 1.26 or less (Tt/K≦1.26).

For example, in the example shown in FIG. 4B, the sum of the length Ta of the first sliding surface 126a and the length Tb of the second sliding surface 126b is the total sliding length Tt (Tt=Ta+Tb). Therefore, in this example, it is sufficient that Ta+Tb≦1.26. Also, in the example shown in FIG. 4C, the sum of the length Tc of the first sliding surface 126c, the length Td of the second sliding surface 126d, and the length Tf of the third sliding surface 126e is the total sliding length Tt (Tt=Tc+Td+Te). Therefore, in this example, it is sufficient that Tc+Td+Te≦1.26.

For the sake of convenience, the configuration in which the ratio T/K≦0.51 based on the aforementioned single sliding length T is referred to as the "first configuration of the main shaft sliding part," and the configuration in which the ratio Tt/K≦1.26 based on the aforementioned total sliding length Tt is referred to as the "second configuration of the main shaft sliding part." Then, only the first configuration may be combined with the characteristic configuration in embodiment 1, or only the second configuration may be combined with the characteristic configuration in embodiment 1. Alternatively, both the first and second configurations may be combined with the characteristic configuration in embodiment 1.

FIG. 4A shows an example in which the first configuration is applied to the spindle sliding portion, and FIGS. 4B and 4C show examples in which both the first and second configurations are applied to the spindle sliding portion, but it goes without saying that the present disclosure is not limited to the configurations shown in FIGS. 4A to 4C. As mentioned above, only the second configuration can be applied to the spindle sliding portion.

In this way, when there are multiple sliding surfaces, if the ratio T/K is 0.51 or less and the ratio Tt/K is 1.26 or less, the wear resistance of the main shaft sliding part, which is derived from the sulfur-based sliding property modifier described below, can be further improved when the sliding area is reduced by using a low viscosity oil (with a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to ^2.5 mm2/s) as the refrigeration oil 180.

In this disclosure, the lower limit of the ratio Tt/K is not particularly limited, but an example of a preferred lower limit is 0.3 or more. Therefore, a preferred range of the ratio Tt/K in this disclosure is within the range of 0.3 to 1.26. A more preferred lower limit of the ratio Tt/K is 0.60, and an even more preferred lower limit is 0.99. Generally, if the ratio Tt/K is 0.3 or more, the sliding area is not excessively reduced even if the sliding surface is divided into multiple surfaces. Therefore, even if a low-viscosity oil is used as the refrigeration oil 180, the wear resistance of the main shaft sliding portion can be suitably achieved by using a sulfur-based sliding property modifier.

In the examples shown in Figures 4A to 4C, the ratio T/K or the ratio Tt/K is described for the main shaft 124 of the crankshaft 120 as the shaft portion, but the present disclosure is not limited to this and is similar to the eccentric shaft 122. In this embodiment 2, as described in the above embodiment 1, the connecting portion between the eccentric shaft 122 and the connecting means 142 is the sliding portion, or in other words, a part of the connecting means 142 that slidably connects the eccentric shaft 122 corresponds to the "eccentric bearing."

Therefore, when the sliding surface of the eccentric shaft 122 with the "eccentric bearing" (the connection portion of the eccentric shaft 122 and the connection means 142) is a single surface, the axial length of the sliding surface is taken as the single sliding length T, or when the sliding surface of the eccentric shaft 122 is divided into multiple surfaces, the axial length of the sliding surface with the smallest axial length is taken as the single sliding length T, and the ratio T/K of the single sliding length T to the outer diameter K of the eccentric shaft 122 should be 0.51 or less. Also, when the sum of the axial lengths of the multiple sliding surfaces of the eccentric shaft 122 is taken as the total sliding length Tt, the ratio Tt/K of the total sliding length Tt to the outer diameter K of the eccentric shaft 122 should be 1.26 or less.

Therefore, in this second embodiment, the refrigerant compressor 100 only needs to satisfy the "first configuration" in which the ratio T/K is 0.51 or less in at least one of the shaft parts, the main shaft 124 and the eccentric shaft 122. Alternatively, it only needs to satisfy the "second configuration" in which the ratio Tt/K is 1.26 or less in at least one of the main shaft 124 and the eccentric shaft 122. Furthermore, it is also possible for at least one of the main shaft 124 and the eccentric shaft 122 to satisfy both the first and second configurations.

Therefore, if the sliding part formed by the connecting part of the eccentric shaft 122 and the connecting means 142 is regarded as the "eccentric shaft sliding part", then any "main shaft sliding part" in the description of the second embodiment can be replaced with the "eccentric shaft sliding part". Furthermore, when the first or second configuration is applied to the eccentric shaft sliding part, it can also be expressed as the "first configuration of the eccentric shaft sliding part" or the "second configuration of the eccentric shaft sliding part".

[Sulfur-based friction modifier]
The refrigeration oil 180 used in the present embodiment 2 is as described in the above embodiment 1, and may be a low-viscosity oil having a kinetic viscosity in the range of 1.0 mm ² /s to 2.5 mm ² /s at 40° C. A more specific configuration of the refrigeration oil 180 in the present disclosure will be described in the embodiment described later.

Here, in the second embodiment, the low-viscosity oil that is the refrigeration oil 180 contains a sulfur-based sliding property modifier. As described in the first embodiment, in the refrigerant compressor 100 according to the present disclosure, the crankshaft 120 is made of an iron-based material. The specific type of iron-based material is not particularly limited, and examples include metal materials that contain iron as a main component, such as various known cast irons and steel materials. The sulfur-based sliding property modifier may be one that is capable of reacting with such an iron-based material and sulfur.

Therefore, the sliding property modifier in the present embodiment 2 may be sulfur itself, or may be a sulfur compound that contains sulfur and is capable of reacting with iron-based materials. For example, examples of sulfur compounds that can be used as sliding property modifiers include sulfurized olefins, sulfide-based compounds (e.g., dibenzyl (di)sulfide (DBDS), etc.), xanthates, thiadiazoles, thiocarbonates, sulfurized oils and fats, sulfurized esters, dithiocarbamates, and sulfurized terpenes.

The amount of the sulfur-based sliding property modifier contained in the refrigeration oil 180 is not particularly limited. Typically, the sliding property modifier is added to the refrigeration oil 180 so that the amount is 100 ppm or more when converted to elemental sulfur weight (mass). The lower limit of the amount of sliding property modifier added (content), 100 ppm converted to elemental sulfur weight, is greater than the upper limit of the amount generally added of the sulfur-based extreme pressure additive described below.

If the content (addition amount) of the sliding property modifier is less than 100 ppm when converted into elemental sulfur weight, depending on various conditions, when a low-viscosity refrigeration oil 180 is used and the sliding area of the main shaft sliding part is reduced, suitable wear resistance of the main shaft sliding part may not be achieved. In addition, a preferable lower limit of the content of the sulfur-based sliding property modifier can be, for example, 150 ppm or more when converted into elemental sulfur weight. In addition, a preferable upper limit of the content of the sulfur-based sliding property modifier can be, for example, 1000 ppm or less when converted into elemental sulfur weight, and more preferably 500 ppm or less.

The sulfur-based sliding property modifier used in this disclosure may be a compound similar to known sulfur-based extreme pressure additives, but it may be one that is relatively more reactive with the shaft material than known extreme pressure additives, or it may be added to the refrigeration oil 180 in an amount greater than the typical amount (content) of known extreme pressure additives.

In general, extreme pressure additives are compounds that contain active elements such as sulfur, halogen elements, and phosphorus, and they react chemically with the material surfaces (sliding surfaces) that make up the sliding parts to form a coating that suppresses wear, seizure, fusion, etc. of the sliding members. However, it is also well known that compounds that contain sulfur easily react with copper.

In the refrigerant compressor 100, copper wire is used as the winding of the electric element 104. Furthermore, in a freezing/refrigeration device using the refrigerant compressor 100, copper pipes are generally used as the refrigerant piping. As mentioned above, copper is susceptible to corrosion when reacting with compounds containing sulfur, so when using a sulfur-based extreme pressure additive, measures are required to avoid or suppress corrosion of the copper members (or copper-containing members) of the refrigerant compressor 100 or the freezing/refrigeration device and to prevent a decrease in their reliability.

Therefore, in the field of refrigerant compressors 100, it is common knowledge to use special compounds in combination with sulfur-based extreme pressure additives to prevent them from reacting with copper or copper-containing components of the refrigerant compressor 100 or the refrigeration/freezing device. Alternatively, it is also known not to use sulfur-based compounds as additives at all.

In response to this, the inventors of the present invention have conducted extensive research, including experimental verification, and as a result, it has become clear that when a low-viscosity refrigeration oil 180 is used and the sliding area of the main shaft sliding part is reduced so that the aforementioned ratio T/K is 0.51 or less (first configuration of the main shaft sliding part) or the aforementioned ratio Tt/K is 1.26 or less (second configuration of the main shaft sliding part), not only can good wear resistance be achieved, but corrosion of copper members (or copper-containing members) can also be substantially avoided by using a more reactive sulfur-based compound as a sliding property modifier or increasing the amount added (content).

It is widely known in the field of lubricants that friction modifiers and extreme pressure additives are clearly different components.

When the oil film breaks in the sliding section and metallic contact occurs between the sliding parts, the surface layer (e.g., the oxide layer) is removed from the contact area of each sliding surface, creating new metallic protrusions. These metallic protrusions on the sliding surfaces may fuse to each other. The sliding property modifier forms a coating (anti-wear film) in place of the removed surface layer. This makes it possible to prevent the metallic protrusions from fusing to each other, thereby effectively suppressing wear in the sliding section.

In contrast, extreme pressure additives quickly form a coating (extreme pressure film, EP film) to replace the removed surface layer. This EP film is formed more strongly on the sliding surface than the anti-wear film formed by sliding property modifiers. This is because extreme pressure additives are intended to suppress wear in sliding parts that are in a lubricated state where the contact pressure between the sliding surfaces is relatively high and the oil film is prone to rupture, that is, in an "extreme pressure state."

Extreme pressure additives are generally used as additives added to refrigeration oil 180 to suppress wear in the sliding parts of refrigerant compressor 100. In contrast, it is not common to add sliding property modifiers, which form a film slower than extreme pressure additives. However, in this second embodiment, it is expected that adding a sulfur-based sliding property modifier will cause a film to form at a moderate rate on the main shaft sliding parts, making it easier for sulfur to become localized (unevenly distributed) on the main shaft sliding parts.

As a result, even if a higher concentration of sulfur-based compounds (sulfur-based extreme pressure additives) is added, not only is it possible to achieve good sliding properties in the main shaft sliding parts, but it is also believed that corrosion of the copper members (or copper-containing members) in the refrigerant compressor 100 or freezing/refrigeration equipment is suppressed.

In this way, in the second embodiment, in the refrigerant compressor 100 using a low-viscosity oil having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to ^{2.5 mm2} /s as the refrigerating machine oil 180, at least one of the configuration for increasing the average piston speed (V), the configuration for setting the ratio S/D, and the configuration for setting the ratios L1/D and L2/L1 described in the first embodiment is applied, and further, at least in the main shaft sliding portion, a first configuration in which the ratio T/K of the single sliding length T to the outer diameter K of the shaft portion is 0.51 or less, or a second configuration (or both the first and second configurations) in which the ratio Tt/K of the total sliding length Tt to the outer diameter K of the shaft portion is 1.26 or less, and a configuration in which a sulfur-based sliding property modifier is used is applied.

By applying at least one of the configurations described in the first embodiment, even when a low-viscosity oil is used as the refrigeration oil 180, leakage of the refrigerant gas 181 can be effectively suppressed, and an even better coefficient of performance (COP) can be achieved in the refrigerant compressor 100. Furthermore, by applying the first or second configuration and the configuration using a sulfur-based sliding property modifier described in the second embodiment, the main shaft sliding part can be effectively lubricated and wear of the main shaft sliding part can be effectively suppressed. As a result, the reliability of the refrigerant compressor 100 can be further improved.

Moreover, in the present disclosure, if the refrigerant compressor 100 is configured to be inverter-driven, the electric element 104 may be operated at a low rotation speed (low-speed operation) or at a high rotation speed (high-speed operation). In particular, in the present disclosure, the electric element 104 may be operated at a low rotation speed with an operating frequency of 16 r/s or more and 35 r/s or less. In general, in low-speed operation, the oil supply capacity of the oil supply mechanism 125 provided on the crankshaft 120 decreases, and the amount of refrigeration oil 180 supplied to each sliding part tends to decrease.

In this disclosure, as explained in the first embodiment, even during low-speed operation, the average piston speed can be increased and the viscous force of the oil film can be increased. This makes it possible to suppress an increase in sliding loss in the cylinder sliding parts, and also to suppress leakage of the refrigerant gas 181. This makes it possible to achieve a good coefficient of performance (COP).

On the other hand, by applying the first or second configuration to the main shaft sliding portion, the sliding area between the main shaft 124 and the main bearing 134 becomes relatively small, but good wear resistance can be achieved even if the supply amount of refrigeration oil 180 decreases. Therefore, an increase in sliding loss in the main shaft sliding portion can also be suppressed, and a good coefficient of performance (COP) can be achieved.

Therefore, by combining and applying the configuration described in the first embodiment and the configuration described in the second embodiment to the refrigerant compressor 100, it is possible to achieve an even better coefficient of performance (COP).

In addition, if both the first and second configurations are applied to the refrigerant compressor 100 in the main shaft sliding portion, the lubrication condition in the main shaft sliding portion can be further improved. Therefore, the coefficient of performance (COP) can be further improved.

(Embodiment 3)
The hermetic refrigerant compressor according to the third embodiment has the same basic configuration as the hermetic refrigerant compressor according to the first embodiment, but has a more characteristic thrust bearing. Since the basic configuration of the hermetic refrigerant compressor according to the third embodiment is the same as the configuration shown in Fig. 1 of the first embodiment, detailed description thereof will be omitted.

[Thrust bearing]
An example of a specific configuration of the thrust bearing in the second embodiment will be specifically described with reference to Fig. 5 and Fig. 6. Fig. 5 and Fig. 6 are each a schematic diagram of a part of the cross section of the refrigerant compressor 100 shown in Fig. 1. Fig. 5 is a schematic diagram showing an example of distances P and Q set in the thrust bearing provided in the refrigerant compressor 100 and a load (main shaft load) applied to the main shaft sliding portion. Fig. 6 is a schematic diagram showing an example of a main configuration of the thrust bearing.

As shown in FIG. 1, in the refrigerant compressor 100, the main bearing 134 has a tubular or cylindrical shape that extends vertically relative to the main body of the cylinder block 130, which has a "lateral" extension within the sealed container 102. The main body of the main bearing 134 extends below the cylinder block 130. And, as explained in the first embodiment, the tubular extension 137 extends above the cylinder block 130. Therefore, the main body of the main bearing 134 and the tubular extension 137 form a single tubular or cylindrical structure.

The inner circumferential surface of the main bearing 134 is the sliding surface as described above. Therefore, as shown in FIG. 5, the upper edge of the inner circumferential surface of the main bearing 134 is the upper end 138 of the sliding surface, and the lower edge of the main bearing 134 is the lower end 139 of the sliding surface. In this third embodiment, since the main bearing 134 has a tubular extension 137 on the upper side, the upper end 138 of the sliding surface corresponds to the upper edge of the inner circumferential surface of the tubular extension 137. In other words, the tubular extension 137 can be said to be an "extension" that extends the main bearing 134 upward.

By providing such a tubular extension 137, when the upper limit of the distance Q described below is specified, the overall length of the main bearing 134 can be increased without increasing the overall height of the refrigerant compressor 100, and the position of the crankshaft 120 inserted in the main bearing 134 during operation can be improved.

As shown in FIG. 6, the inner surface of the upper end of the tubular extension 137 may be chamfered or otherwise processed. In this case, the inner edge of the chamfered portion of the inner surface of the tubular extension 137 becomes the upper end 138 of the sliding surface of the main bearing 134. Note that if the inner surface of the upper end of the tubular extension 137 is not chamfered or otherwise processed, the upper edge of the inner surface of the tubular extension 137 becomes the upper end 138 of the sliding surface of the main bearing 134.

As shown in FIG. 5, when the distance between the axis of the compression chamber 133 and the lower end 139 of the sliding surface of the main bearing 134 is defined as "distance P," and the distance between the axis of the compression chamber 133 and the upper end 138 of the sliding surface of the main bearing 134 is defined as "distance Q," in the refrigerant compressor 100 according to the present disclosure, even if a thrust bearing such as a thrust ball bearing 210 is provided, when distance P is within the range of 38 mm to 51 mm, distance Q is 16 mm or less.

In the third embodiment, the refrigerant compressor 100 is provided with a thrust bearing on the thrust surface 136 of the main bearing 134. The specific configuration of the thrust bearing is not particularly limited and may be any type of rolling bearing, but in the third embodiment, a thrust ball bearing 210 is used as shown in FIG. 1, FIG. 5 or FIG. 6. As shown in FIG. 6, the thrust ball bearing 210 includes a lower race 206 located on the thrust surface 136, an upper race 202 located opposite the lower race 206, and a plurality of balls 204 as rolling elements that rollably contact the lower race 206 and the upper race 202. Note that a vibration reducing member such as an elastic member may be provided between the thrust surface 136 of the main bearing 134 and the lower race 206.

The thrust ball bearing 210 is disposed on the outer periphery of the tubular extension 137, and the balls 204 are housed in a retainer 205. The upper race 202 and the lower race 206 are, for example, annular metal flat plates, and are disposed parallel to each other. The upper race 202 and the lower race 206 may be provided with arc-shaped grooves.

In the example configuration shown in FIG. 6, the lower race 206, balls 204, and upper race 202 are stacked in this order on top of the thrust surface 136 while in contact with each other, and the flange portion 128 of the crankshaft 120 is seated on the top surface of the upper race 202. This forms the thrust ball bearing 210.

The thrust ball bearing 210 is a rolling bearing in which the balls 204 roll in point contact with the upper race 202 and the lower race 206. Therefore, the thrust ball bearing 210 can support a vertical load while rotating the main shaft 124 with little friction. Note that while the thrust ball bearing 210 is a "ball bearing" that uses balls 204 as the rolling elements, it may also be a "roller bearing" that uses rollers as the rolling elements, or it may be another rolling bearing.

As a result, the bearing function of the sliding bearing is changed to a rolling bearing called the thrust ball bearing 210, thereby reducing losses and effectively increasing the efficiency of the refrigerant compressor 100. However, normally, providing a thrust bearing such as the thrust ball bearing 210 increases the overall height of the refrigerant compressor 100.

In contrast, in the refrigerant compressor 100 disclosed herein, the distances P and Q based on the axis of the compression chamber 133 are set so that when the distance P is within the range of 38 mm to 51 mm, the distance Q is 16 mm or less.

Generally, in order to reduce the sliding loss in the spindle 124, it is possible to adopt a configuration that reduces the friction coefficient in the sliding part of the spindle and/or a configuration that reduces the load on the spindle 124 (spindle load F2). Furthermore, in order to reduce the spindle load F2, it is possible to adopt a configuration that reduces the distance Q and/or a configuration that increases the distance P.

However, to increase the distance P, it is necessary to increase (raise) the overall height of the refrigerant compressor 100. If the overall height is increased in this way, it will be necessary to expand the engine room (machine room) of the refrigeration/freezer in which the refrigerant compressor 100 is installed, which will ultimately lead to a reduction in the internal volume of the refrigeration/freezer. Therefore, in order to reduce the main shaft load F2, it is expected that the distance Q will be reduced without changing the distance P.

However, if one were to simply try to reduce the distance Q, it would be necessary to adopt a method of thinning the support part of the cylinder block 130 or reducing the thickness of the flange part 128 to 4 mm or less, i.e., a method of thinning a specific member (or a part of it) (thinning method).

However, adopting such a thinning method ends up causing deformation of other components. Specifically, thinning the support portion reduces the rigidity of the cylinder block 130, making the main bearing 134 more susceptible to deformation, and thinning the flange portion 128 increases the inclination of the eccentric shaft 122. In particular, the increase in the inclination of the eccentric shaft 122 caused by thinning the flange portion 128 was not anticipated in the past.

In this way, by reducing the distance Q using a thinning technique, the efficiency of the refrigerant compressor 100 can be increased, but there is a risk that the reliability of the refrigerant compressor 100 will decrease due to deformation of certain components.

In contrast, in this third embodiment, as a result of experimental verification, it was uniquely discovered that by setting the upper limit of the distance Q to a predetermined value, i.e., 16 mm or less, it is possible to achieve both high efficiency and good reliability without employing thinning techniques.

Specifically, it has become clear that when the distance Q is reduced, the slight tilt (inclination angle) of the eccentric shaft 122 that occurs during operation of the refrigerant compressor 100 is involved not only in the reliability of the refrigerant compressor 100 but also in its high efficiency. In other words, this finding means that the change in the distance Q and the tilt of the eccentric shaft 122 are important factors in reducing the main shaft load F2 and achieving high efficiency and good reliability of the refrigerant compressor 100. Therefore, after extensive investigations, the inventors have come to the conclusion that it is important to set the upper limit of the distance Q to 16 mm or less.

In this third embodiment, when distance P is set within the range of 38 mm to 51 mm, distance Q is set to 16 mm or less, or distance Q may be set within the range of 12 mm to 16 mm (i.e., 12 mm as an example of a lower limit value). Therefore, there is no need to increase (make high) the overall height of refrigerant compressor 100. This not only makes it possible to achieve high efficiency while maintaining the good quality (especially reliability) of refrigerant compressor 100, but also ensures sufficient internal volume of the refrigeration/freezing device, since there is no need to expand the engine room (machine room) of the refrigeration/freezing device.

In this way, in the third embodiment, in a refrigerant compressor 100 equipped with a thrust bearing, when the distance P that affects the overall height is set within a predetermined range, the upper limit of the distance Q between the axis of the compression chamber 133 and the upper end 138 of the sliding surface of the main bearing 134 is set to 16 mm. This makes it possible to avoid an increase in the overall height without excessively thinning the flange portion 128, which contributes to the stability of the eccentric shaft 122, and also reduces the load on the main shaft 124 without applying special treatment to the sliding surface.

As a result, it is possible to achieve even higher efficiency without increasing the overall height of the refrigerant compressor 100. Moreover, because the flange portion 128 is not made excessively thin, it is possible to achieve high efficiency as well as good reliability.

Here, in order to reduce the sliding loss of the main shaft sliding part, in addition to reducing the distance Q, it is also possible to reduce the friction coefficient of the main shaft sliding part. If one simply wants to reduce the friction coefficient, it is conceivable to reduce the viscosity of the refrigeration oil 180 as much as possible.

As described in the first embodiment, the refrigerant compressor 100 according to the present disclosure uses, as the refrigerant oil 180, an extremely low-viscosity oil having a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 mm ² /s. Therefore, the friction coefficient can be reduced simply by using this low-viscosity oil as the refrigerant oil 180. Furthermore, as described in the third embodiment, the main shaft load F2 can be further reduced by setting the distance Q of the refrigerant compressor 100 to 16 mm or less. This reduces the sliding loss of the main shaft sliding portion.

Therefore, by combining the configuration described in the third embodiment with the configuration described in the first embodiment, it is possible to achieve high efficiency and good reliability of the refrigerant compressor 100 not only in the cylinder sliding parts but also in the main shaft sliding parts. This makes it possible to further improve the coefficient of performance (COP) of the refrigerant compressor 100.

Furthermore, in this third embodiment, there is no need to particularly limit the diameter of piston 140, i.e., piston diameter (D), or the inner diameter of compression chamber 133 into which piston 140 is inserted. If distance Q is set to 16 mm or less when distance P is within the range of 38 mm to 51 mm, not only is there no need to make flange portion 128 excessively thin, but there is also no need to substantially regulate piston diameter (D) or the inner diameter of compression chamber 133.

As explained in the first embodiment, by setting the ratio S/D of the stroke amount (S) to the piston diameter (D) within the range of 0.78 to 1.00, the viscous force (F1) of the oil film can be increased in the cylinder sliding portion (see formula (1) above). Setting this ratio S/D leads to a reduction in the piston diameter (D), but in the third embodiment, the piston diameter (D) does not need to be substantially regulated. Therefore, the configuration explained in the third embodiment has the advantage of being easily applicable to the configuration explained in the first embodiment.

Furthermore, according to the third embodiment, by setting the distance Q to 16 mm or less and reducing the spindle load F2, it is possible to easily form a good oil film during low-speed operation even if a low-viscosity oil is used as the refrigeration oil 180. As explained in the first embodiment, in particular in this disclosure, low-speed operation with an operating frequency of 16 r/s or more and 35 r/s or less may be performed.

Therefore, when the configuration described in this third embodiment is applied to the configuration described in the first embodiment, it can be fully applied even during low-speed operation, making it possible to effectively suppress or avoid wear or seizure in the main shaft sliding parts. Therefore, even when the refrigerant compressor 100 is operating at low speed, the configuration described in this third embodiment can be easily applied to the configuration described in the first embodiment.

Furthermore, as explained in the second embodiment, by combining the configuration explained in the first embodiment with the configuration explained in the second embodiment, it is possible to realize an even better coefficient of performance (COP) in the refrigerant compressor 100. And, by combining the configuration explained in the third embodiment with the configuration explained in the first embodiment, it is possible to make the coefficient of performance (COP) even better. Therefore, by combining the configurations explained in the first embodiment, the second embodiment, and the third embodiment, it is possible to achieve a favorable synergistic effect in terms of the effect of realizing a good coefficient of performance (COP).

In this third embodiment, as shown in Figs. 1 and 5, an eccentric shaft 122 is provided on the upper part (upper end) of the main shaft 124, and a piston 140 is connected to this eccentric shaft 122 via a connecting means 142, and this piston 140 is inserted so as to be able to reciprocate within a compression chamber 133 arranged in the horizontal direction. That is, in this third embodiment, the piston 140 and the compression chamber 133 are located at the upper part within the refrigerant compressor 100. However, the configuration of the refrigerant compressor 100 according to the present disclosure is not limited to this.

For example, although not shown, an eccentric shaft 122 may be provided at the bottom (lower end) of the main shaft 124, so that the piston 140 and the compression chamber 133 are located at the bottom of the refrigerant compressor 100. In this case, the distance P is the distance between the axis of the compression chamber 133 and the upper end of the sliding surface, and the distance Q is the distance between the axis of the compression chamber 133 and the lower end of the sliding surface.

Alternatively, in the third embodiment, as shown in FIG. 1, the crankshaft 120 extends in the "vertical direction" (up and down direction) of the refrigerant compressor 100, and therefore the main shaft 124 and the eccentric shaft 122 also extend in the vertical direction. However, the configuration of the refrigerant compressor 100 according to the present disclosure is not limited to this, and for example, the crankshaft 120 may extend in the "horizontal direction" (direction perpendicular to the vertical direction), and the piston 140 and the compression chamber 133 may be biased to one side of the horizontal direction rather than the vertical direction within the refrigerant compressor 100. In this case, both ends of the sliding surface that are the basis for the distances P and Q are located in the horizontal direction rather than the vertical direction.

Therefore, in this disclosure, the end of the sliding surface of the main bearing 134 on the compression chamber 133 (or eccentric shaft 122) side is defined as the first end, and the opposite end is defined as the second end. Therefore, distance P can be defined as the distance between the axis of the compression chamber 133 and the second end of the sliding surface of the main bearing 134, and distance Q can be defined as the distance between the axis of the compression chamber 133 and the first end of the sliding surface of the main bearing 134. In this third embodiment (the example shown in FIG. 1 or FIG. 5), upper end 138 of the sliding surface is the first end, and lower end 139 of the sliding surface is the second end.

The refrigeration oil 180 used in this third embodiment may be the low-viscosity oil described above, but as described in the fourth embodiment below, the low-viscosity oil may contain a high molecular weight component (the preferred oil described below). Such a preferred oil allows for a better oil film to be formed in the sliding parts. This further improves the effects and advantages obtained by the configuration described in this third embodiment, and further the effects and advantages obtained by the configuration described in the first embodiment (and also the effects and advantages when the configuration described in the second embodiment is applied).

(Embodiment 4)
The hermetic refrigerant compressor according to the fourth embodiment has a basic configuration similar to at least any one of the hermetic refrigerant compressors according to the first to third embodiments described above, but has a further characteristic configuration in terms of the low viscosity oil used as the refrigerating machine oil 180. Note that in the fourth embodiment, a specific configuration example of the refrigerating machine oil 180 applicable to the refrigerant compressor 100 described in any one of the first to third embodiments will be described. Therefore, a specific description of the refrigerant compressor 100 will be omitted.

As described above, the refrigeration oil 180 according to the present disclosure is not particularly limited as long as it is a low-viscosity oil having a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 ^{mm 2} /s. As a representative refrigeration oil 180, for example, at least one oily substance selected from the group consisting of mineral oil, alkylbenzene oil, and ester oil can be suitably used. As described above, a representative oily substance can be mineral oil.

These oily substances may be used alone or in appropriate combination of two or more. A combination of two or more oily substances here does not only include, for example, a combination of two or more different oily substances that fall under the mineral oil category, but also includes, for example, a combination of one or more oily substances that fall under the mineral oil category and one or more oily substances that fall under the alkylbenzene oil category (or one or more oily substances that fall under the ester oil category).

The refrigeration oil 180 according to the present disclosure may contain various known additives in addition to the oily substance. As such additives, various additives known in the field of refrigeration oil 180 can be suitably used, but representative examples include sliding property modifiers, extreme pressure additives, oiliness agents, antioxidants, acid scavengers, metal deactivators, antifoaming agents, corrosion inhibitors, and dispersants.

In particular, in the second embodiment, a sulfur-based sliding property modifier is added to the low-viscosity oil used as the refrigeration oil 180, but a known extreme pressure additive may also be added. Specific extreme pressure additives that can be used suitably include known ones, and are not particularly limited, but examples include phosphorus-based compounds such as phosphate esters, halogenated compounds such as chlorinated hydrocarbons or fluorinated hydrocarbons, etc. Only one type of these extreme pressure additives may be added to the low-viscosity oil (oily substance), or two or more types may be added in appropriate combination.

Among these extreme pressure additives, phosphorus-based compounds can be preferably used. Representative phosphorus-based compounds include tricresyl phosphate (TCP), tributyl phosphate (TBP), and triphenyl phosphate (TPP), with TCP being more preferably used. In particular, in the configuration described in the second embodiment, by adding a phosphorus-based extreme pressure additive to the refrigeration oil 180 in addition to a sulfur-based sliding property modifier, it is possible to achieve good wear reduction in the main shaft sliding parts.

The amount of extreme pressure additive added to the low-viscosity oil is not particularly limited. For example, if the refrigeration oil 180 (oily substance) is a low-polarity substance such as mineral oil or alkylbenzene oil, the amount can be in the range of 0.5 to 8.0 mass%, or can be in the range of 1 to 3 mass%, when the total mass of the low-viscosity oil is taken as 100 mass%.

In addition, a sliding property modifier, an extreme pressure additive, or other additives may be added to the refrigeration oil 180 used in the configuration described in the first or third embodiment. These additives can be added to the refrigeration oil 180 according to the present disclosure as long as they do not interfere with the effects obtained by the configuration described in the first to third embodiments and as long as the effects derived from the additives can be obtained.

In other words, the refrigeration oil 180 used in the refrigerant compressor 100 according to the present disclosure may be composed of an oily substance having a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 mm ² /s. When two or more types of oily substances are used to constitute the "oil composition", the kinetic viscosity of the oil composition at 40° C. may be within the above range. Furthermore, the refrigeration oil 180 according to the present disclosure may be an "oil composition" containing, in addition to one or more types of oily substances, a sulfur-based sliding property modifier (or other sliding property modifier), a phosphorus-based extreme pressure additive (or other extreme pressure additive), or other additives.

Therefore, in the present disclosure, the oil composition used as the refrigeration oil 180 can be said to be a "low viscosity oil" having a kinetic viscosity at 40° C. in the range of 1.0 mm ^{2 /} s to 2.5 mm ² /s.

Furthermore, the oily substance used in the refrigeration oil 180 according to the present disclosure may have a molecular weight within a predetermined range. Specifically, the number average molecular weight Mn of the oily substance used as the refrigeration oil 180 may be 150 to 400. Furthermore, the weight average molecular weight Mw (or mass average molecular weight) of the oily substance may be 150 to 400, and may be within the range of 200 to 300.

Furthermore, the polydispersity index (PDI) of the oily substance, i.e., the ratio of number average molecular weight Mn to weight average molecular weight Mw, Mw/Mn, may be within the range of 1.0 to 1.1. There are no particular limitations on the method for measuring the oily substance and its molecular weight (number average molecular weight Mn and weight average molecular weight Mw) described below, but in the present disclosure, one example is standard polystyrene conversion using the GPC (Gel Permeation Chromatography) method.

Generally, when the viscosity of the oily substance used as the refrigeration oil 180 is reduced, the molecules of the oily substance are reduced in molecular weight. When such a low molecular weight oily substance comes into contact with the resin material present inside the refrigerant compressor 100, there is concern that it may cause a "deterioration in extractability," in which the components (extractable components) contained in the resin material may become more easily extracted.

In contrast, if the polydispersity Mw/Mn of the oily substance is at least within the range of 1.0 to 1.1, the variation in the molecular weight of the oily substance is small, and the molecular weight of the oily substance is prevented from becoming excessively small. Therefore, it is possible to significantly reduce the "deterioration of extractability" of the refrigeration oil 180, that is, the possibility that the refrigeration oil 180 will extract extractable components from the resin material used in the refrigerant compressor 100.

The "deterioration of extractability" in refrigeration oil 180 means that extractable components extracted from the resin material are mixed into the low-viscosity oil used as refrigeration oil 180, which may result in a deterioration in the quality of refrigeration oil 180. If the quality of refrigeration oil 180 deteriorates, not only may good lubrication not be achieved at the sliding parts of the cylinder or the sliding parts of the main shaft, but it may also be that a good oil film will not be formed at the sliding parts of the cylinder. In this case, it may not be possible to effectively suppress leakage of refrigerant gas 181 from between piston 140 and cylinder 132. Therefore, the molecular weight and polydispersity of the low-viscosity oil (its main component, the oily substance) used as refrigeration oil 180 may be within the above-mentioned ranges.

Furthermore, as explained in the third embodiment, the refrigeration oil 180 according to the present disclosure may contain a component having a relatively large molecular weight, i.e., a high molecular weight component, in addition to the oily substance. Therefore, the refrigeration oil 180 according to the present disclosure may be, for example, an "oil composition" that contains an oily substance having a molecular weight in the above-mentioned predetermined range as a main component, and further contains a high molecular weight component.

In this disclosure, the oil composition (low viscosity oil) used as the refrigeration oil 180 is not limited to a composition containing high molecular weight components, so in the following explanation, the oil composition containing high molecular weight components will be referred to as the "preferred oil" for the sake of convenience.

The high molecular weight component contained in the suitable oil may have a weight average molecular weight Mw (mass average molecular weight) of 500 or more. In addition, the content of the high molecular weight component may be 0.5 mass% or more when the total mass of the oil composition used as the refrigeration oil 180 is taken as 100 mass%.

The suitable oil used as the refrigeration oil 180 in this embodiment 4 may be one that originally contains high molecular weight components, or may be one to which an oily substance corresponding to a high molecular weight component is added so as to be 0.5 mass% or more. An example of the former is mineral oil. When preparing (producing) a suitable oil by refining unrefined or roughly refined raw mineral oil, the refining conditions or refining method of the raw oil may be adjusted so that 0.5 mass% or more of high molecular weight components remain. An example of the latter is one in which mineral oil, alkylbenzene oil, or polyalkylene glycol oil is used as the "main component" of the suitable oil, and an oily substance that becomes a high molecular weight component is added to this main component as an "additive component".

The molecular weight and polydispersity of the oily substance that is the main component of the suitable oil may be within the ranges described above. If the molecular weight and polydispersity of the suitable oil are within these ranges, and the oil contains 0.5 mass% or more of high molecular weight components, it will be possible to form a suitable oil film on the main shaft sliding portion, particularly when the distance Q is set to 16 mm or less in the thrust bearing configuration described in the third embodiment.

The upper limit of the content of high molecular weight components is not particularly limited, so long as it does not affect the function or effect of the preferred oil. However, typical examples of the upper limit of the content of high molecular weight components include 7.0% by mass or less, 6.0% by mass or less, and 5.0% by mass.

Although it depends on various conditions such as the specific configuration of the refrigerant compressor 100 or the specific composition of the refrigeration oil 180, if the content of the high molecular weight component exceeds 7.0 mass %, it may affect the viscosity of the oil (oil composition) used as the refrigeration oil 180. In this case, the kinetic viscosity of the oil (oil composition) at 40° C. may exceed the range of 1.0 mm ² /s to 2.5 mm ² /s. Therefore, it may not be possible to obtain an improvement in the coefficient of performance (COP) commensurate with the content of the high molecular weight component.

Furthermore, the reason why the coefficient of performance (COP) of the refrigerant compressor 100 is improved by the suitable oil containing a high molecular weight component is believed to be that, even if the suitable oil has a low viscosity (kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to 2.5 ^mm2 /s), the high molecular weight component contributes to the formation of a good oil film in the sliding parts, according to the results of experimental verification. Therefore, when a suitable oil containing a high molecular weight component is used as the refrigeration oil 180, it is believed that an even better oil film is formed not only in the main shaft sliding parts but also in the cylinder sliding parts. Therefore, not only can good lubrication be achieved in these sliding parts, but leakage of the refrigerant gas 181 in the cylinder sliding parts can also be further suppressed.

When the suitable oil is a main component to which a high molecular weight component has been added, the specific material or type of the high molecular weight component is not particularly limited, and it may be an oily substance with a weight average molecular weight Mw of 500 or more. For example, when the main component is mineral oil, the high molecular weight component may be the same mineral oil, alkylbenzene oil, polyalkylene glycol oil, or another oily substance.

Furthermore, when the suitable oil is an oil composition in which a high molecular weight component is added as an additive component to an oily substance as the main component, for example, one type of oily substance may be used as the main component and one type of oily substance different from the main component may be used as the high molecular weight component. Alternatively, two or more types of oily substances may be used as the main component and one type of oily substance as the high molecular weight component, or one type of oily substance may be used as the main component and two or more types of oily substances as the high molecular weight component. Alternatively, two or more types of mixtures of oily substances in which a high molecular weight component has been added to the main component may be further mixed.

As described above, the refrigeration oil 180 according to the fourth embodiment may contain a sliding property modifier (e.g., sulfur or a sulfur-containing compound), an extreme pressure additive (e.g., a phosphorus-containing compound), or other known additives in addition to one or more types of oily substances that are the main components. Here, when a suitable oil containing a high molecular weight component is used as the refrigeration oil 180, an oily agent may be added as an additive. When the suitable oil contains an oily agent, an oil film of the suitable oil is more easily formed on the sliding surface of the sliding part. This makes it possible to more effectively achieve low friction in the sliding parts, and also to suppress leakage of the refrigerant gas 181 in the cylinder sliding parts.

The specific type of oily agent is not particularly limited, but representative examples include higher fatty acids, higher alcohols, esters (ester compounds), ethers, amines, amides, metal soaps, etc. These oily agents may be used alone or in combination of two or more types. The amount of oily agent added is not particularly limited, but when the total mass of the suitable oil (oil composition) is taken as 100 mass%, it may be within the range of 0.01 to 1 mass%, for example.

Among the oily agents mentioned above, ester compounds are particularly representative. The ester compounds may be any compound having an ester structure formed by reacting an alcohol with a carboxylic acid. The alcohol may be a monohydric alcohol or a polyhydric alcohol having two or more hydric groups. Similarly, the carboxylic acid may be a monocarboxylic acid, a dicarboxylic acid, or a tricarboxylic acid (which may have four or more carboxy groups). Generally, commercially available ester oily agents can be used.

If the suitable oil is an oil composition containing an oiliness agent, the oil film forming ability can be further improved. In other words, since the suitable oil contains high molecular weight components, the high molecular weight components are present on the sliding surface of the sliding part (spindle sliding part or cylinder sliding part, etc.), which is thought to enable the formation of a good oil film. Furthermore, if the suitable oil contains an oiliness agent, the oiliness agent is adsorbed to the sliding surface, which is thought to further facilitate the formation of an oil film by the suitable oil (oil composition).

In particular, if the oily agent is an ester compound, the oily agent will have an ester bond. Therefore, the polarity resulting from this ester bond makes it easier for the oil film of the suitable oil (oil composition) to adhere more closely to the sliding parts (improving the adhesion of the oil film). This further improves the oil film forming ability of the suitable oil, so that the friction coefficient can be further reduced and low friction of the sliding parts can be more suitably achieved. Furthermore, in the cylinder sliding parts, a good oil film is more likely to be formed between the piston 140 and the cylinder 132, so leakage of the refrigerant gas 181 can be more suitably suppressed.

When the refrigeration oil 180 according to the fourth embodiment is the preferred oil described above, as described above, it may contain a sulfur-based sliding property modifier or a phosphorus-based extreme pressure additive as an additive. By containing these additives in the preferred oil, not only can the action and effect obtained from each additive be imparted to the preferred oil, but a synergistic effect of each additive can be expected, so that not only can the sliding performance of each sliding part be further improved, but leakage of the refrigerant gas 181 in the cylinder sliding part can also be effectively suppressed.

On the other hand, the suitable oil may not contain at least one of the sulfur-based sliding property modifier, phosphorus-based extreme pressure additive, and ester-based oil agent, or all of them, depending on the specific configuration or various conditions of the refrigerant compressor 100. In other words, the suitable oil only needs to contain suitable additives as necessary, and the specific additives are not limited to the sulfur-based sliding property modifier, phosphorus-based extreme pressure additive, ester-based oil agent, etc., mentioned above. In addition, a low-viscosity oil that does not contain high molecular weight components may be used as the refrigeration oil 180.

(Embodiment 5)
In the fifth embodiment, an example of a freezing/refrigerating device including the refrigerant compressor 100 described in the first to fourth embodiments will be specifically described with reference to FIG.

The refrigerant compressor 100 according to the present disclosure can be widely and suitably used in various devices (freezing/refrigeration devices) having a refrigeration cycle or a configuration substantially equivalent thereto. Specific examples include, but are not limited to, refrigerators (domestic refrigerators, commercial refrigerators), ice makers, showcases, dehumidifiers, heat pump water heaters, heat pump washer-dryers, vending machines, air conditioners, and air compressors. In this second embodiment, the basic configuration of a freezing/refrigeration device is described using an item storage device shown in FIG. 7 as an application example of the refrigerant compressor 100 according to the present disclosure.

As shown in FIG. 7, the freezing/refrigeration device of this embodiment 5 includes a main body 301, a partition wall 304, and a refrigerant circuit 305. The main body 301 is composed of an insulated box and a door, and the box has one side open, and the door opens and closes the opening of the box. The inside of the main body 301 is partitioned by the partition wall 304 into an item storage space 302 and a machine room 303. A blower (not shown) is provided in the storage space 302. The inside of the main body 301 may be partitioned into spaces other than the storage space 302 and the machine room 303.

The refrigerant circuit 305 is configured to cool the inside of the storage space 302, and includes the refrigerant compressor 100 described in each of the above embodiments, a radiator 307, a pressure reducing device 308, and a heat absorber 309, which are connected in a ring shape by piping. In other words, the refrigerant circuit 305 is an example of a refrigeration cycle using the refrigerant compressor 100 according to the present disclosure.

As mentioned above, refrigerant gas 181, such as R600a, is sealed inside refrigerant compressor 100 (sealed container 102), and this refrigerant gas 181 is sealed at a relatively low temperature so that the pressure is equivalent to the low pressure side of the freezing/refrigeration device. Although there are no particular limitations on the specific type of refrigerant gas 181, it is preferable to use a hydrocarbon-based refrigerant gas with a low global warming potential, such as R600a.

The heat absorber 309 of the refrigerant circuit 305 is disposed in the storage space 302. The cooling heat of the heat absorber 309 is stirred by a blower (not shown) so as to circulate within the storage space 302, as shown by the dashed arrows in FIG. 7. This cools the inside of the storage space 302.

In this way, the refrigeration/storage device according to the fifth embodiment is equipped with the refrigerant compressor 100 according to any one of the first to fourth embodiments. The refrigerant compressor 100 can further improve its coefficient of performance (COP) when using a low-viscosity oil having a kinetic viscosity at 40° C. in the range of 1.0 mm ² /s to 2.5 mm ² /s as the refrigeration oil 180. Therefore, the refrigeration/storage device equipped with such a refrigerant compressor 100 can reduce its power consumption.

The present invention will be described in more detail based on reference examples, examples, and conventional examples, but the present invention is not limited thereto. Those skilled in the art may make various changes, modifications, and alterations without departing from the scope of the present invention.

(Reference example)
In a conventional refrigerant compressor, a total of four types of mineral oils with different kinetic viscosities at 40°C (1.8 ^mm2 /s, ^2.5 ^mm2 /s, 3.3 mm2/s, and 5.0 ^mm2 /s) were used as refrigerating machine oil 180 (the kinetic viscosity of refrigerating machine oil 180 was changed) and the amount of refrigerant gas 181 leaking from between piston 140 and cylinder 132 (refrigerant leakage amount) was evaluated when the operating frequency was set to 17 r/s. The results are shown in the graph of FIG.

The amount of refrigerant leakage in the reference example was measured (evaluated) as follows. In the target refrigerant compressor (see FIG. 1), the suction hole (not shown in FIG. 1) was modified to be blocked, and a container (refrigerant container) containing a certain amount of refrigerant gas 181 was connected to the discharge hole (not shown in FIG. 1) side to prepare a refrigerant leakage measurement system. In this measurement system, since the suction hole is blocked, refrigerant gas 181 is introduced from the refrigerant container through the gap between piston 140 and cylinder 132 into the hermetic refrigerant compressor. Therefore, in this measurement system, the pressure reduction in the refrigerant container can be regarded as the amount of refrigerant leakage. In this measurement system, the refrigerant compressor was operated at an arbitrary frequency to evaluate the amount of refrigerant leakage when four types of mineral oil with different kinetic viscosities were used.

In the graph of Fig. 8, the horizontal axis represents the kinetic viscosity (unit: ^mm2 /s) of the refrigerating machine oil 180 at 40°C, and the vertical axis represents the amount of refrigerant leakage (unit: %). The evaluation standard for the amount of refrigerant leakage in Fig. 8 is set to 100% when a refrigerating machine oil 180 with a kinetic viscosity of 5.0 ^mm2 /s at 40°C is used.

As is clear from the results in Figure 8, with conventional refrigerant compressors, the amount of refrigerant leakage clearly increases as the viscosity decreases, especially during low-speed operation (17 r/s).

(Evaluation method of coefficient of performance)
The results of the reference example revealed that in a conventional refrigerant compressor, when operating at low speed, the amount of refrigerant leakage increases significantly if the kinetic viscosity at 40°C is 2.5 ^mm2 /s or less. Therefore, an evaluation was conducted to see how the coefficient of performance (COP) changes in the refrigerant compressor 100 to which the configuration for increasing the average piston speed (V), the configuration for setting the ratio S/D, and the configuration for setting the ratios L1/D and L2/L1 described in the first embodiment are applied. The above configurations applied to the refrigerant compressor 100 are shown in Table 1.

The coefficient of performance (COP) of the refrigerant compressor 100 of the embodiment or the conventional refrigerant compressor was calculated as the ratio of refrigeration capacity to energy consumption (input) (refrigeration capacity/input). The evaluation standard of the coefficient of performance (COP) was set to 100% when using refrigeration oil 180 with a kinetic viscosity of 5.0 ^mm2 /s at 40°C.

(Example)
The refrigerant compressor 100 of the present disclosure, which has the configuration shown in Table 1 (average piston speed, ratio S/D, ratio L1/D, ratio L2/L1), was used, and a total of seven types of mineral oils with different kinetic viscosities at 40°C (1.8 ^mm2/ s, ^2.2 mm2/s, 2.3 mm2 ^/ s, ^2.5 mm2/s, ^2.7 ^mm2 /s, 3.3 mm2/s, and 5.0 ^mm2 /s) were used as the refrigerating oil 180 (by changing the kinetic viscosity of the refrigerating oil 180) to evaluate its coefficient of performance (COP).

The results when the operating frequency was 27 r/s are shown in Fig. 9, and the results when the operating frequency was 17 r/s are shown in Fig. 10. In the graphs of Fig. 9 and Fig. 10, the results of the example are indicated by circle symbols. In both graphs, the horizontal axis represents the kinetic viscosity (unit: ^mm2 /s) of the refrigeration oil 180 at 40°C, and the vertical axis represents the coefficient of performance (COP).

(Conventional example)
A conventional refrigerant compressor, that is, a refrigerant compressor having the same configuration as the example except that the configuration shown in Table 1 is not applied, was used, and a total of five types of mineral oils having different kinetic viscosities at 40°C (1.8 ^mm2 /s, ^{2.3 mm2/} s, ^{2.7 mm2} /s, 3.3 ^mm2 /s, and 5.0 ^mm2 /s) were used as the refrigeration oil 180 (the kinetic viscosity of the refrigeration oil 180 was changed) to evaluate the coefficient of performance (COP). The results when the operating frequency was 27 r/s are shown in Figure 9, and the results when the operating frequency was 17 r/s are shown in Figure 10. In the graphs of Figures 9 and 10, the results of the conventional example are indicated by a cross symbol.

(Comparison between the embodiment and the conventional example)
9, when the operating frequency is 27 r/s, in both the embodiment and the conventional example, the coefficient of performance (COP) improves when the kinetic viscosity of the refrigeration oil 180 is reduced. This is because the input power to the refrigerant compressor is reduced as the viscous resistance of the refrigeration oil 180 is reduced.

10, when the operating frequency is lower at 17 r/s (low-speed operation), in the conventional example, the coefficient of performance (COP) decreases when the kinetic viscosity (40°C) of the refrigeration oil 180 becomes 2.5 ^mm2 /s or less. This is because, as described above, when the kinetic viscosity of the refrigeration oil 180 becomes 2.5 ^mm2 /s or less, the amount of leakage of the refrigerant gas 181 from between the piston 140 and the cylinder 132 increases, and the refrigeration capacity decreases.

On the other hand, in the examples, the coefficient of performance (COP) is improved even when the kinetic viscosity (40°C) of the refrigeration oil 180 is 2.5 ^mm2 /s or less. Therefore, it can be seen that in the refrigerant compressor 100 (sealed refrigerant compressor according to the present disclosure) to which the configuration shown in Table 1 is applied, when a low-viscosity oil having a kinetic viscosity (40°C) in the range of 1.0 ^mm2 /s to ^{2.5 mm2} /s is used as the refrigeration oil 180, a good coefficient of performance (COP) can be achieved even during low-speed operation.

Also, as shown in Table 1, the average piston speed in the conventional example is 0.31 m/s at 17 r/s, while the average piston speed in the embodiment is 0.34 m/s at 17 r/s. As described above, in the embodiment, a good coefficient of performance (COP) can be achieved even when low viscosity oil is used as the refrigeration oil 180, but in the conventional example, a good coefficient of performance (COP) cannot be achieved when low viscosity oil is used as the refrigeration oil 180. Therefore, it can be seen that a good coefficient of performance (COP) can be achieved if the average piston speed is at least 0.31 m/s within the range of 16 r/s to 35 r/s, which is defined as the low speed operating frequency in this disclosure.

(Additional Note)
Based on the above description of the embodiments, the following techniques are disclosed in this specification.

(Technique 1)
A hermetic refrigerant compressor comprising: a hermetic container; refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2/s to 2.5 ^mm2 /s; a cylinder block housed in the hermetic container and forming a compression chamber; and a piston inserted and reciprocatable within the compression chamber, wherein the average reciprocating speed of the piston is set to exceed 0.31 m/s when the operating frequency is 16 r/s or more and 35 r/s or less.

(Technique 2)
The hermetic refrigerant compressor according to claim 1, wherein a ratio S/D of a stroke amount (S) of the piston to a piston diameter (D) is within a range of 0.78 to 1.00.

(Technique 3)
The hermetic refrigerant compressor according to Technology 1 or Technology 2, wherein, when a length of an area where the piston seals the inside of the compression chamber by its reciprocating motion is defined as a seal length (L2), a ratio L1/D of the piston total length (L1) to the piston diameter (D) is within a range of 0.8 to 1.0, and a ratio L2/L1 of the seal length (L2) to the piston total length (L1) is within a range of 0.9 to 1.0.

(Technique 4)
A hermetic refrigerant compressor comprising: a hermetic container; refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2/s to 2.5 ^mm2 /s; a cylinder block housed in the hermetic container and forming a compression chamber; and a piston inserted in the compression chamber so as to be able to reciprocate, wherein a ratio S/D of a reciprocating stroke amount (S) of the piston to a piston diameter (D) is in the range of 0.78 to 1.00.

(Technique 5)
The hermetic refrigerant compressor according to technique 4, wherein, when a length of an area where the piston seals the inside of the compression chamber by its reciprocating motion is defined as a seal length (L2), a ratio L2/L1 of the seal length (L2) to a total piston length (L1) is within a range of 0.9 to 1.0.

(Technique 6)
a cylinder block that is housed in the sealed container and forms a compression chamber; and a piston that is inserted reciprocally within the compression chamber, wherein a ratio ^L1 /D of the piston overall length ( ^L1 ) to the piston diameter (D) is within a range of 0.8 to 1.0, and wherein, when a length of an area where the piston seals the compression chamber by its reciprocating motion is defined as a seal length (L2), a ratio L2/L1 of the seal length (L2) to the piston overall length (L1) is within a range of 0.9 to 1.0.

(Technique 7)
The compression element includes, as a shaft portion, a crankshaft having a main shaft and an eccentric shaft, and, as a bearing portion supporting the shaft portion, a main bearing supporting the main shaft and an eccentric bearing supporting the eccentric shaft, a sliding surface of the main shaft with the main bearing is divided into a plurality of surfaces, and when a sum of axial lengths of the plurality of sliding surfaces is a total sliding length Tt, a ratio Tt/K of the total sliding length Tt to an outer diameter K of the main shaft is 1.26 or less.

(Technique 8)
The compression element includes, as a shaft portion, a crankshaft having a main shaft and an eccentric shaft, and, as a bearing portion supporting the shaft portion, a main bearing supporting the main shaft and an eccentric bearing supporting the eccentric shaft, a sliding surface of the main shaft with the main bearing is a single surface or is divided into a plurality of surfaces, and when the sliding surface is a single surface, when an axial length of the sliding surface is defined as a single sliding length T, or when the sliding surface is divided into a plurality of surfaces, when the axial length of a sliding surface having a smallest axial length is defined as a single sliding length T, a ratio T/K of the single sliding length T to an outer diameter K of the main shaft is 0.51 or less, and further, the refrigerating machine oil contains sulfur or a compound having sulfur as a sliding property modifier.

(Technique 9)
The compression element further includes a crankshaft having a main shaft and an eccentric shaft, a main bearing supporting the main shaft, and a thrust bearing provided on a thrust surface of the main bearing, wherein an end portion of a sliding surface of the main bearing on the compression chamber side is a first end and an end portion on the opposite side is a second end, a distance between an axis of the compression chamber and a second end of the sliding surface of the main bearing is P, and a distance between the axis of the compression chamber and the first end of the sliding surface of the main bearing is Q, wherein when the distance P is within a range of 38 mm to 51 mm, the distance Q is 16 mm or less.

(Technique 10)
A method for operating a hermetic refrigerant compressor comprising: a hermetic container; refrigeration oil stored in the hermetic container and having a kinetic viscosity at 40°C in the range of 1.0 ^mm2 /s to ^{2.5 mm2} /s; a cylinder block housed in the hermetic container and forming a compression chamber; and a piston inserted and reciprocatable within the compression chamber, the method comprising the steps of: when the operating frequency is 16 r/s or more and 35 r/s or less, causing the average reciprocating speed of the piston to exceed 0.31 m/s.

(Technique 11)
A freezing and refrigeration device comprising a refrigerant circuit including the hermetic refrigerant compressor according to any one of Techniques 1 to 9, or the hermetic refrigerant compressor in which the operating method according to Technique 9 is performed, a radiator, a pressure reducing device, and a heat absorber, and connecting these in a ring shape by piping.

The present invention is not limited to the above-described embodiment, and various modifications are possible within the scope of the claims. The technical scope of the present invention also includes embodiments obtained by appropriately combining the technical means disclosed in different embodiments or multiple modified examples.

Furthermore, many improvements and other embodiments of the present invention will be apparent to those skilled in the art from the above description. Therefore, the above description should be interpreted as merely illustrative and is provided for the purpose of teaching those skilled in the art the best mode for carrying out the present invention. Details of the structure and/or function of the present invention may be substantially changed without departing from the spirit of the present invention.

As described above, according to the present invention, a hermetic refrigerant compressor can achieve an even better coefficient of performance (COP) by using a refrigeration oil with a lower viscosity. Therefore, the present invention can be widely applied to various devices that use a refrigeration cycle.

100: Hermetic refrigerant compressor 102: Hermetic container 104: Electric element 106: Compression element 108: Compressor body 120: Crankshaft 122: Eccentric shaft 124: Main shaft 125: Oil supply mechanism 126: Sliding surface 126a: First sliding surface 126b: Second sliding surface 126c: First sliding surface 126d: Second sliding surface 126e: Third sliding surface 127: Non-sliding outer peripheral surface (non-sliding surface)
127a: First non-sliding outer peripheral surface (non-sliding surface)
127b: second non-sliding outer peripheral surface (non-sliding surface)
128: flange portion 130: cylinder block 132: cylinder 133: compression chamber 134: main bearing 136: thrust surface 137: tubular extension portion 138: upper end (first end) of sliding surface
139: Lower end of sliding surface (second end)
140: Piston 142: Connection means 150: Stator 152: Rotor 180: Refrigerating machine oil 181: Refrigerant gas 190: Suspension spring 202: Upper race 204: Ball (rolling element)
205: Cage 206: Lower race 210: Thrust ball bearing (thrust bearing)
240: conventional piston 301: main body 302: storage space 303: machine room 304: partition wall 305: refrigerant circuit 307: radiator 308: pressure reducing device 309: heat sink

Claims

A sealed container;
a refrigeration oil having a kinetic viscosity at 40° C. in the range of 1.0 mm 2 /s to 2.5 mm 2 /s stored in the sealed container;
The compressor comprises a cylinder block that is housed in the sealed container and forms a compression chamber, and a piston that is inserted into the compression chamber so as to be capable of reciprocating motion,
When the operating frequency is 16 r/s or more and 35 r/s or less, the average speed at which the piston reciprocates is set to be greater than 0.31 m/s.
Hermetic refrigerant compressor.
The ratio S/D of the reciprocating stroke amount (S) of the piston to the piston diameter (D) is within a range of 0.78 to 1.00.
2. The hermetic refrigerant compressor according to claim 1.
When the length of the area where the piston seals the inside of the compression chamber by its reciprocating motion is defined as a seal length (L2),
A ratio L1/D of the piston overall length (L1) to the piston diameter (D) is within a range of 0.8 to 1.0,
The ratio L2/L1 of the seal length (L2) to the piston total length (L1) is within a range of 0.9 to 1.0.
3. The hermetic refrigerant compressor according to claim 1 or 2.
A sealed container;
a refrigeration oil having a kinetic viscosity at 40° C. in the range of 1.0 mm 2 /s to 2.5 mm 2 /s stored in the sealed container;
The compressor comprises a cylinder block that is housed in the sealed container and forms a compression chamber, and a piston that is inserted into the compression chamber so as to be capable of reciprocating motion,
The ratio S/D of the reciprocating stroke amount (S) of the piston to the piston diameter (D) is within a range of 0.78 to 1.00.
Hermetic refrigerant compressor.
When the length of the area where the piston seals the inside of the compression chamber by its reciprocating motion is defined as a seal length (L2),
The ratio L2/L1 of the seal length (L2) to the piston total length (L1) is within a range of 0.9 to 1.0.
5. The hermetic refrigerant compressor according to claim 4.
A sealed container;
a refrigeration oil having a kinetic viscosity at 40° C. in the range of 1.0 mm 2 /s to 2.5 mm 2 /s stored in the sealed container;
The compressor comprises a cylinder block that is housed in the sealed container and forms a compression chamber, and a piston that is inserted into the compression chamber so as to be capable of reciprocating motion,
A ratio L1/D of the piston overall length (L1) to the piston diameter (D) is within a range of 0.8 to 1.0,
When the length of the area where the piston seals the inside of the compression chamber by its reciprocating motion is defined as a seal length (L2),
The ratio L2/L1 of the seal length (L2) to the piston total length (L1) is within a range of 0.9 to 1.0.
Hermetic refrigerant compressor.
The compression element includes a crankshaft having a main shaft and an eccentric shaft as a shaft portion, and a main bearing supporting the main shaft and an eccentric bearing supporting the eccentric shaft as a bearing portion supporting the shaft portion,
A sliding surface of the spindle with the main bearing is divided into a plurality of surfaces, and when a total axial length of the plurality of sliding surfaces is defined as a total sliding length Tt, a ratio Tt/K of the total sliding length Tt to an outer diameter K of the spindle is 1.26 or less.
7. A hermetic refrigerant compressor according to claim 1, 4 or 6.
The compression element includes a crankshaft having a main shaft and an eccentric shaft as a shaft portion, and a main bearing supporting the main shaft and an eccentric bearing supporting the eccentric shaft as a bearing portion supporting the shaft portion,
The sliding surface of the main shaft with the main bearing is a single surface or is divided into a plurality of surfaces,
When the sliding surface is a single surface, the axial length of the sliding surface is defined as a single sliding length T, or when the sliding surface is divided into a plurality of surfaces, the axial length of the sliding surface having the shortest axial length is defined as the single sliding length T, and a ratio T/K of the single sliding length T to an outer diameter K of the spindle is 0.51 or less;
Furthermore, the refrigerating machine oil contains sulfur or a compound having sulfur as a sliding property improver.
7. The hermetic refrigerant compressor of claim 1, 4 or 6.
The compression element further includes a crankshaft having a main shaft and an eccentric shaft, a main bearing supporting the main shaft, and a thrust bearing provided on a thrust surface of the main bearing,
When an end portion of the sliding surface of the main bearing on the compression chamber side is defined as a first end and an end portion on the opposite side is defined as a second end, a distance between the axis of the compression chamber and the second end of the sliding surface of the main bearing is defined as P, and a distance between the axis of the compression chamber and the first end of the sliding surface of the main bearing is defined as Q,
When the distance P is within a range of 38 mm to 51 mm, the distance Q is 16 mm or less.
7. The hermetic refrigerant compressor of claim 1, 4 or 6.
A sealed container;
a refrigeration oil having a kinetic viscosity at 40° C. in the range of 1.0 mm 2 /s to 2.5 mm 2 /s stored in the sealed container;
A hermetic refrigerant compressor including a cylinder block accommodated in the hermetic container and forming a compression chamber, and a piston inserted in the compression chamber so as to be capable of reciprocating motion,
When the operating frequency is 16 r/s or more and 35 r/s or less, the average speed at which the piston reciprocates exceeds 0.31 m/s.
A method for operating a hermetic refrigerant compressor.
A refrigerant circuit including the hermetic refrigerant compressor according to claim 1, 4, or 6, a radiator, a pressure reducing device, and a heat sink, the refrigerant circuit being connected in a ring shape by piping.
Refrigeration and freezing equipment.