WO2011110019A1

WO2011110019A1 - Rotary compressor

Info

Publication number: WO2011110019A1
Application number: PCT/CN2010/077128
Authority: WO
Inventors: 小津政雄; 李华明
Original assignee: 广东美芝制冷设备有限公司
Priority date: 2010-03-10
Filing date: 2010-09-20
Publication date: 2011-09-15
Also published as: CN102192149B; CN102192149A

Abstract

A rotary compressor is provided with a motor portion (22) and a compression mechanism portion (21) within a hermetic housing (1). The compression mechanism portion (21) includes more than one cylinder (23). In the cylinder (23) is provided a cylinder compression chamber (24) in which a piston (28) is accommodated. A leading end of a vane is in contact with the periphery of the piston and divides the cylinder compression chamber into a high-pressure chamber (31a) and a low-pressure chamber (31b). A crankshaft (27) engages with the piston to drive it. A main bearing (25) and a secondary bearing (26) are used to support the crankshaft and are installed on the cylinder. The main bearing is connected with a main shaft (34a) of the crankshaft, and the secondary bearing (26) is connected with a secondary shaft (34b) of the crankshaft. In the position of at least one of two ends (36) of the piston adjacent to one's end face is provided a circumferential groove (37) opening toward the inside of the piston. A flexible thin wall (38) is provided on the outer side of the circumferential groove. The compressor can adjust the clearance forming between two ends and the main and secondary bearings, respectively, by the pressure difference between the inside and outside of the piston, and thereby reduces leakage loss and increases energy consumption.

Description

Rotary compressor

Technical field

The present invention relates to a rotary compressor.

Background technique

From the viewpoint of preventing global warming, the efficiency of a rotary compressor mounted on an air conditioner or a refrigeration system is a major issue. Since the piston height gap is formed between the piston and the bearing of the rotary compressor, the high pressure gas leaks from the inside of the piston to the cylinder compression chamber to re-expand, which has a great influence on the efficiency of the rotary compressor.

technical problem

The object of the present invention is to provide a simple and reasonable structure, flexible operation, low production cost, and effective reduction of high-pressure gas leakage loss from the inside of a rotary compressor piston, and can greatly improve the energy efficiency of the compressor and have a wide application range. Rotary compressors to overcome the deficiencies in the prior art.

Technical solution

A rotary compressor designed according to the purpose, a motor part and a compression mechanism part are arranged in the sealed casing, the compression mechanism part comprises more than one cylinder, a cylinder compression chamber is arranged in the cylinder, and the piston is accommodated in the cylinder compression chamber, and the piston is arranged The apex of the piece is connected to the outer circumference of the piston and divides the cylinder compression chamber into a high pressure chamber and a low pressure chamber. The crankshaft is connected to the piston and drives the piston, and the main bearing and the auxiliary bearing for supporting the crankshaft and mounted on the cylinder, the main bearing and the main bearing The main shaft of the crankshaft is connected, and the auxiliary bearing is connected to the auxiliary shaft of the crankshaft, and is characterized in that at least one of the two end portions of the piston is provided with a circumferential groove at a position facing the end surface thereof, and the opening of the circumferential groove faces the inside of the piston The outer side of the circumferential groove is provided with an elastic thin wall; the gap formed between the two ends and the main bearing and the sub-bearing is adjusted by the pressure difference between the inside and the outside of the piston.

The compression mechanism portion includes a first cylinder and a second cylinder, the middle plate is disposed between the first cylinder and the second cylinder, the first piston is disposed in the first cylinder, and the second piston is disposed in the second cylinder, the circumferential groove Provided on at least one of the four ends of the first piston and the second piston; adjusting the ends of the first piston and the second piston by a pressure difference between the inside and the outside of the first piston and/or the second piston A gap formed between any of the main bearing, the sub-bearing, and the intermediate plate.

The shaft diameter of the main shaft of the crankshaft is larger than the shaft diameter of the counter shaft, and the width of the end of the piston toward the sub-bearing is larger than the width of the end portion toward the main bearing.

The piston is provided with an oil hole, one end of the oil hole is opened at the end surface of the piston, and the other end of the oil hole is opened in the circumferential groove.

An annular groove is disposed on an end surface of the piston, and the annular groove communicates with the oil hole.

A recess for increasing the amount of elastic deformation of the thin wall is provided in the circumferential groove.

The recess is disposed on a side of the circumferential groove that faces the thin wall.

A lubricating film is disposed on an end surface of the main bearing and/or the sub-bearing bearing toward the piston.

A lubricating film is disposed on an end surface of the main bearing, the sub-bearing, and/or the intermediate plate toward the piston side.

The rotary compressor is a horizontal rotary compressor or a vertical rotary compressor, and the rotary compressor constitutes a refrigeration cycle with a condenser, an expansion device, and an evaporator.

The invention provides a circumferential groove along the two ends of the piston and from the inside of the piston, so that the two ends of the thinned piston can be elastically deformed according to the pressure difference between the inside of the piston and the compression chamber of the cylinder, thereby optimizing the piston. The height gap, therefore, can reduce the leakage of high pressure gas, prevent the efficiency of the compressor from decreasing, and make the rotary compressor easier to start.

Beneficial effect

The invention has the advantages of simple and reasonable structure, flexible operation, low production cost, effective reduction of high-pressure gas leakage loss from the piston of the rotary compressor, and greatly improved energy efficiency and wide application range of the compressor.

DRAWINGS

Fig. 1 is a schematic view showing the structure of a first embodiment of the present invention.

Figure 2 is a cross-sectional view taken along line X-X of Figure 1.

Fig. 3 is a schematic enlarged plan view showing a compression mechanism portion in the first embodiment.

Fig. 4 is a graph showing the efficiency of a conventional rotary compressor.

Fig. 5 is a view showing a state of force of a compression mechanism unit in the first embodiment;

Fig. 6 is a view showing the elastic deformation of the thin wall in the first embodiment.

Fig. 7 is a view showing a mode change when the rotation angle θ of the piston in the first embodiment is 90 degrees when the cylinder compression chamber is eccentrically rotated.

Fig. 8 is a view showing a mode change when the rotation angle θ of the piston in the first embodiment is 190 degrees when the cylinder compression chamber is eccentrically rotated.

Fig. 9 is a view showing a mode change when the rotation angle θ of the piston in the first embodiment is 360 degrees when the cylinder compression chamber is eccentrically rotated.

Fig. 10 is a graph showing the height clearance of the piston in the first embodiment.

Figure 11 is a schematic view showing the structure of the first embodiment of the piston in the first embodiment.

Figure 12 is a schematic view showing the structure of a second embodiment of the piston in the first embodiment.

Figure 13 is a schematic view showing the structure of a third embodiment of the piston in the first embodiment.

Figure 14 is a schematic view showing the structure of a fourth embodiment of the piston in the first embodiment.

Figure 15 is a schematic view showing the structure of a second embodiment of the present invention.

Figure 16 is a cross-sectional structural view showing a piston in Embodiment 3 of the present invention.

Figure 17 is a top plan view of Figure 16.

Figure 18 is a partially enlarged schematic view of Figure 16;

Figure 19 is a schematic view showing the structure of another embodiment of the piston in Embodiment 3 of the present invention.

Figure 20 is a top plan view of Figure 19.

Figure 21 is a partially enlarged schematic view of Figure 19.

Figure 22 is a partial cross-sectional structural view showing a fourth embodiment of the present invention.

Figure 23 is a partial cross-sectional structural view showing a fifth embodiment of the present invention.

In the figure: 1 is a rotary compressor, 2 is a sealed casing, 3 is a discharge pipe, 5 is an outdoor heat exchanger, 7 is an expansion valve, 8 is a liquid reservoir, 9 is a suction pipe, 10 is a discharge hole, 11 For the spout valve, 13 is the shaft hole, 14 is the eccentric shaft oil hole, 15 is the oil pool, 17 is the eccentric shaft, 18a is the main bearing mounting surface, 18b is the auxiliary bearing mounting surface, 21 is the compression mechanism part, 22 is the motor Department, 23 is the cylinder, 24 is the cylinder compression chamber, 25 is the main bearing, 26 is the auxiliary bearing, 27 is the crankshaft, 28 is the piston, 29 is the sliding piece, 31a is the high pressure chamber, 31b is the low pressure chamber, 32 is the lubricating film, 34a is the main shaft, 34b is the countershaft, 36 is the end of the piston, 37 is the circumferential groove, 38 is the thin wall, 39 is the concave position, 40 For the cylinder screw, 41 is the oil hole, 42 is the annular groove, 50 is the double cylinder rotary compressor, 51 is the double cylinder compression mechanism part, 52a is the first cylinder, 52b is the second cylinder, 53 is the middle plate, 54a is The first piston, 54b is a second piston, 55 is a two-cylinder bearing, 56 is a main bearing, 57 is a sub-bearing, and 60 is a sealed housing.

BEST MODE FOR CARRYING OUT THE INVENTION

The invention is further described below in conjunction with the drawings and embodiments.

The horizontal rotary compressor will be described below as an example.

Example 1

Referring to Fig. 1, a rotary compressor 1 is composed of a compression mechanism portion 21 attached to a cylindrical sealed casing 2 and a motor portion 22 disposed at an upper portion thereof. The compression mechanism portion 21 includes a cylinder 23, a piston 28 that eccentrically rotates in the cylinder compression chamber 24, a slide 29 that abuts against the piston 28 and reciprocates, a crankshaft 27 that drives the piston 28, and a main bearing 25 that lubricates the support crankshaft 27 and The sub-bearings 26 are formed, and the above-described respective components are assembled by five sets of cylinder screws 40. The main bearing 25 is provided with a discharge device including a discharge port 10 and a discharge valve 11.

The assembled compression mechanism portion 21 is fixed to the inner wall of the sealed casing 2 by the outer circumference of the cylinder 23. The oil is injected from the discharge pipe 3 and hoarded in the oil pool 15 at the bottom of the casing 2.

In the system in which the rotary compressor 1 is mounted, the low-pressure gas injected from the suction pipe 9 is compressed in the cylinder compression chamber 24 to become a high-pressure gas, and the high-pressure gas is discharged from the discharge port 10, and is discharged to the seal through the discharge valve 11. The inside of the housing 2. Therefore, the internal pressure of the sealed casing 2 is a high pressure side equivalent to the discharge pressure. Thereafter, the high-pressure gas discharged from the discharge pipe 3 starts from the outdoor heat exchanger 5, flows into the accumulator 8 from the indoor heat exchanger 6 via the expansion valve 7, and finally flows into the suction pipe 9, and is sucked into the cylinder to be compressed. In the cavity 24.

Referring to Figs. 2 to 3, a cylinder compression chamber 24 disposed at a central portion of the cylinder 23 fixed to the inner wall of the seal housing 2 is divided into a high pressure chamber 31a and a low pressure chamber 31b by a piston 28 and a slide 29.

The piston 28, which is driven by the eccentric shaft 17 of the crankshaft 27, rotates clockwise several times per second as indicated by the arrow, and is synchronized with the rotation speed of the motor portion 22 along the inner wall of the cylinder compression chamber 24 to perform about 50 per second. ~60 times of eccentric rotation. The slider 29, which is in contact with the outer wall of the piston 28, reciprocates in synchronism with the eccentric rotation of the piston 28. Therefore, the low pressure chamber 31b and the high pressure chamber 31a change the volume at a high speed.

The circular sub-bearing mounting surface 18b located outside the cylinder compression chamber 24 is subjected to precision grinding with a flatness of about 2 μm. As shown in FIG. 3, the sub-bearing 26 is firmly attached to the sub-bearing mounting surface 18b by the cylinder screw 40, and the main bearing 25 is firmly attached to the main bearing mounting surface 18a of the upper surface of the cylinder 23.

Here, the dimension formed between the main bearing mounting surface 18a of the cylinder 23 and the sub-bearing mounting surface 18b is taken as the cylinder height dimension Hc, and the dimension between the upper and lower end faces of the cylindrical piston 28 is similarly regarded as a piston. The height dimension Hr, which is the sum of the gaps formed between the upper and lower bearing mounting faces of the cylinder 23 and the upper and lower end faces of the piston 28, is referred to as the piston height gap Δr. There is Δr=Hc-Hr.

One half of the piston height clearance Δr, that is, 1/2 Δr, becomes a sliding gap in which the upper and lower ends of the piston 28 slide, respectively. In the rotary compressor 1, the piston height gap Δr or the sliding gap between the upper and lower ends of the piston 28 occupies the longest gap communicating with the cylinder compression chamber 24, and is the most important factor determining the rotational compression efficiency. Therefore, the piston height gap Δr is determined by the height dimension of the selected fitting cylinder and the piston. Generally, the error range is 1 to 2 μm, and the combination is determined.

The components constituting the compression mechanism unit 21: the machining accuracy of the cylinder 23, the piston 28, the main bearing 25, and the sub-bearing 26, the deformation of the compression mechanism portion 21 when assembled by the cylinder screw 40, and the deformation of the seal housing 2 when mounted, are all correct The piston height clearance Δr has a large influence. The machining accuracy of the components can be managed one by one; although the amount of deformation of the compression mechanism portion 21 at the time of assembly can be tested one by one, it is difficult to manage the dispersion.

The components constituting the compression mechanism portion 21 are deformed into: (1) when the main bearing 25 and the sub-bearing 26 are connected to the cylinder 23 by the cylinder screw 40, the three types of components produced are deformed, that is, the main bearing 25, the vice Deformation of each of the bearing 26 and the cylinder 23; (2) deformation of the above-mentioned three components caused by welding and fixing the outer circumference of the compression mechanism portion 21 to the inner wall of the sealing casing 2; (3) deformation of the crankshaft 27 caused by the compression pressure The resulting main bearing 25 and the sub-bearing 26 are deformed; (4) the main bearing 25 and the sub-bearing 26 are generated by the pressure difference between the internal pressure of the sealed casing 2 and the internal pressure of the cylinder compression chamber 24 corresponding to the discharge pressure. The deformation. Moreover, it is also necessary to consider the dispersion of the cylinder, piston height dimension and flatness. Here, the flatness refers to the flatness of the upper and lower end faces of the piston and the cylinder at the same time.

In the future, these deformations and dispersions will be collectively referred to as "bearing deformation", and the amount of deformation will be collectively referred to as "bearing deformation amount".

In the deformation of the bearing, the deformation affecting the piston height gap Δr, particularly the deterioration of the flatness of the main bearing 25 and the sub-bearing 26, is not uneven in a narrow range, and is characterized in that it is generally bowl-shaped or undulating. Therefore, as the angle of rotation of the piston 28 changes, the value of the piston height gap Δr also changes. Where the piston height clearance Δr becomes large, the high-pressure gas leaks from the pressure corresponding to the inside of the sealed casing 2, that is, the inside of the piston on the high-pressure side, to the cylinder compression chamber, thereby lowering the compression efficiency.

An example of this problem will be described. Fig. 4 is a graph showing the relationship between the magnitude of the piston height clearance Δr and the efficiency in a conventional air-conditioner rotary compressor having a cylinder displacement of about 20 cc. The piston height gap Δr used herein is a difference in height between the cylinder and the piston member, that is, the value of the above-described selection fitting is different from the actual piston height gap Δr after the assembly of the compression mechanism portion is completed.

The horizontal axis represents the piston height clearance Δr, the unit is μm; the vertical axis represents power consumption, the unit is KW/h; the unit of cooling capacity is KW/h; energy efficiency COP=cooling/power consumption. For the data of Fig. 4, various motors are common, so the difference in energy efficiency COP can be considered as the efficiency of the compression mechanism portion.

As shown in Fig. 4, the cooling capacity and power consumption are closely related to the change of Δr. As a result, when the piston height gap Δr is 15 μm, the energy efficiency COP is the highest value; the piston height gap Δr is in the range of 14 to 17 μm, and the energy efficiency COP is acceptable.

When the piston height clearance Δr is smaller than 14 μm, the piston height clearance Δr becomes small, so that the sliding friction loss generated between the upper and lower ends of the piston and the ends of the two bearings becomes large. The power consumption has increased. Further, when the piston height clearance Δr is small, the startup failure of the compressor, the reduction in the thickness of the lubricating oil film, and the occurrence of wear are caused, so that the compressor may malfunction.

When the piston height gap Δr is larger than 17 μm, the high-pressure gas that leaks from the inside of the piston and leaks into the low-pressure chamber of the cylinder compression chamber via the upper and lower ends of the piston re-expands, thereby reducing the amount of cooling. Similarly, the inside of the piston leaks into the high-pressure chamber under compression, and the high-pressure chamber at this time is an intermediate pressure, and the high-pressure gas expands here to increase the power consumption. Thus, the high-pressure gas leaking from the inside of the piston into the compression chamber of the cylinder not only lowers the amount of cooling, but also increases the power consumption, thereby causing a problem that the energy efficiency COP is greatly reduced.

In the first embodiment of the present invention, a method for solving the problem is disclosed. As shown in Fig. 3, a circumferential groove 37 is provided between the inner wall and the outer wall at a position close to the end faces of the upper and lower end portions 36 of the piston 28, and elasticity is formed in the upper and lower end portions 36 of the piston. Deformed thin wall 38. Therefore, the upper and lower end faces of the thin wall 38 become the sliding faces of the piston 28 with respect to the main bearing 25 and the sub-bearing 26.

At the time of operation of the compressor, the inside of the piston 28 is brought to a high pressure side corresponding to the pressure of the seal housing 2 by the action of the axial hole 13 penetrating the center of the crankshaft 27 and the eccentric shaft oil hole 14 communicating therewith. By sufficiently supplying oil from the oil pool 15, the outer wall of the eccentric shaft 17 and the inner wall of the piston 28 can be lubricated, and sufficient oil can be retained in the inner space of the piston 28.

Referring to Figure 5, a detailed view of the cylinder compression chamber 24 is shown. The gap between the upper end portion of the piston 28 and the end portion of the main bearing 25 is regarded as Δ1, and the gap between the lower end portion of the piston 28 and the end portion of the sub-bearing 26 is regarded as Δ2. Δ1 + Δ2 = piston height gap Δr. When the compressor is in operation, the piston 28 is eccentrically rotated, causing the bearing of the compression mechanism portion 21 to be deformed, so that Δ1 and Δ2 are relatively independent, and each of them is constantly changed in size.

Fig. 5 is a pressure distribution diagram of the cavity 38 acting on the thin wall 38 in a state where the cylinder compression chamber 24 is divided into the low pressure chamber 31b and the high pressure chamber 31a by the action of the outer circumference of the piston 28 and the slider 29 (not shown).

The pressure of the low pressure chamber 31b is often equal to the suction pressure Ps, and the pressure of the high pressure chamber 31a changes between the suction pressure Ps and the high pressure Pd corresponding to the discharge pressure by the rotation angle of the piston 28, but Fig. 5 shows the high pressure chamber 31a. The discharge pressure Pd is made. The pressure acting on the circumferential groove 37 opening inside the piston 28 is often equivalent to the discharge pressure Pd or the pressure Pd in the sealed casing 2.

The rigidity of the thin wall 38 is sufficiently small, and it is elastically deformed in accordance with the pressure difference acting on the upper and lower sides thereof. FIG. 6 shows a deformed form of the thin wall 38. The two thin walls 38 exhibit a gentle curved curve, which is regarded as C1 and C2, respectively, with the gap between the main bearing 25 and the sub-bearing 26. Therefore, the difference between Δ1 and C1 and the difference between Δ2 and C2 are the amounts of elastic deformation of the respective thin walls 38, respectively. If the design of the two thin walls 38 above and below the piston 28 is the same, the amount of elastic deformation of the piston 28 at the same vertical position in the vertical direction is the same.

The amount of elastic deformation of the thin wall 38 is a variable, and is determined by the thickness and shape design of the thin wall 38, the material elastic modulus of the piston 28, the suction pressure Ps inside the cylinder compression chamber 24, and the magnitude of the discharge pressure Pd. In order to improve the wear resistance, the piston is usually subjected to heat treatment to increase its hardness. Therefore, the elasticity of the material is good.

The low pressure chamber 31b is often the suction pressure Ps, and the amount of elastic deformation of the thin wall 38 becomes maximum, and this mode is referred to as mode 1. Therefore, Mode 1 is such that the pressure difference between the inside and the outside of the piston 28 becomes large, and the gas leaks under the maximum condition, so that the gas leakage is minimized.

When the pressure of the high pressure chamber 31a corresponds to the discharge pressure Pd, the thin wall 38 is not elastically deformed, and this mode is referred to as mode 0. Therefore, in mode 0, neither C1 nor C2 change. However, there is no pressure difference between the inside of the piston 28 and the high pressure chamber 31a, so that the phenomenon that gas leaks from the inside of the piston 28 to the high pressure chamber 31a does not occur.

When the pressure of the high pressure chamber 31a is higher than the suction pressure Ps and lower than the discharge pressure Pd, that is, when the intermediate pressure Pm is located, the elastic deformation amount of the thin wall 38 changes between mode 1 and mode 0, and this mode is called For mode 1-0. In the mode 1-0, the pressure of the high pressure chamber 31a is low, and the amount of elastic deformation of the thin wall 38 is increased. As the pressure of the high pressure chamber 31a rises and the amount of elastic deformation decreases, the amount of gas leaking from the inner diameter of the piston 28 can be automatically reduced.

Referring to Figures 7-9, the mode change of the piston 28 as it rotates clockwise in the cylinder compression chamber 24 is shown. θ is the rotation angle of the piston 28 when the center line of the slider 29 is the base point. The range of θ is 0 to 360 degrees. When the turning angle θ is a certain value, the gap between the piston 28 and the inner circumference of the cylinder compression chamber 24 can be minimized.

Therefore, between the slide piece 29, the cylinder compression chamber 24 is divided into a low pressure chamber 31b and a high pressure chamber 31a.

Fig. 7 shows that when θ is 90 degrees, the low pressure chamber 31b is the suction pressure Ps, and the high pressure chamber 31a is the intermediate pressure Pm. Therefore, in the range of θ = 0 to 90 degrees, the elastic deformation of the thin wall 38 of the piston 28 is mode 1; in the range of θ = 90 to 360 degrees, the elastic deformation of the thin wall 38 of the piston 28 is mode 1-0. .

Fig. 8 shows that when θ is 190 degrees, the low pressure chamber 31b is the suction pressure Ps, and the high pressure chamber 31a is the discharge pressure Pd. Therefore, in the range of θ = 0 to 190 degrees, the elastic deformation of the thin wall 38 is mode 1; in the range of θ = 190 to 360 degrees, the deformation of the thin wall 38 is mode 0.

Fig. 9 shows that when θ is 360 degrees, that is, when the piston 28 and the slider 29 are at the top dead center, the entirety of the cylinder compression chamber 24 becomes the low pressure chamber 31b. Since the pressure is the suction pressure Ps, the thin wall 38 is mode 1 all the week.

In this way, the piston 28 can optimize the clearance of the gaps C1 and C2 during the first eccentric rotation, and the deformation mode of the thin wall 38 can be automatically changed as the angle of rotation changes, so that the amount of gas leaking from the inside of the piston 28 is reached. least.

In Fig. 8, when θ = 190 degrees, that is, when the cylinder compression chamber is divided into a low pressure chamber and a high pressure chamber at the outer circumference of the piston 28, the low pressure chamber side is mode 1, and the high pressure chamber side is mode 0, so the thin wall 38 The difference in the amount of elastic deformation is the largest. In the large range on both sides of the above-mentioned turning angle, the thin wall 38 is gently deformed, so that there is no problem.

An example of the method of setting the elastic deformation amount of the thin wall 38 described above will be described with reference to Fig. 10 . The vertical axis indicates the height Hc of the main bearing mounting surface 18a of the cylinder as a reference surface, and relatively indicates the piston height gap Δr formed between the piston height Hr and the piston. On the horizontal axis, rotors having different piston heights Hr of 1 μm are shown: R1 to R5.

Fig. 10 is a graph showing the amount of change in the bearing as the amount of change in the reference surface of the cylinder, which is represented by a bearing deformation curve. Here, there is a deformation of 5 μm in the direction in which the piston height gap Δr increases and the direction in which the piston is increased. That is to say, the bearing deformation is a concavity and convexity centering on the cylinder reference surface, and the bearing deformation amount is ±5 μm. As described above, the amount of deformation of the bearing is the total amount of deformation of the end portions of the main bearing 25, the sub-bearing 26, and the cylinder, which are caused by the assembly of the compression mechanism portion 21.

In the data of the conventional rotary compressor described in FIG. 4, when the piston height gap Δr at the time of stopping the compressor is set to 15 μm, the highest energy efficiency COP can be obtained.

Fig. 10 is a view mainly illustrating the rotor R3 and the piston height clearance Δr = 15 μm.

The Δr of the rotor R3 when the compressor was stopped was set to 15 μm. As shown in the figure, the minimum gap of the rotor R3 was 10 μm and the maximum gap was 20 μm as the bearing deformation amount changed during the operation of the compressor. Thus, as the amount of deformation of the bearing during operation changes, the piston height gap Δr increases or decreases.

In other words, in the conventional rotary compressor, when the total piston clearance Δr is 10 μm, the piston rotates to generate a compression action. Therefore, the amount of high-pressure gas leaking from the inside of the piston exceeds expectations, which is a cause of a decrease in energy efficiency COP.

In order to solve this problem, the thin wall 38 shown in the first embodiment is added to the upper and lower end portions of the piston, and in the design of the rotor R3 in which the amount of elastic deformation is set to 5 μm, that is, the elastic deformation is added, the gap is 10 μm. The thin wall 38a is elastically deformed and disappears. In Fig. 10, it is expressed by "elastic deformation addition".

Here, in the example applied to the air conditioner, when the difference between the discharge pressure Pd and the suction pressure Ps is 2 MPa, and the surface pressure acting on the thin wall 38 is about 20 kg/cm2 in the mode 1, it is less. Under the surface pressure, after the piston 28 is eccentrically rotated, a lubricating film of 1 μm or more is generated on the sliding surfaces of the main bearing 25 and the sub-bearing 26 of the thin wall 38, respectively.

Therefore, as described above, when the bearing deformation amount is 10 μm, the elastic deformation amount of the upper and lower thin walls 38 of the piston is larger than this, for example, when the respective deformation amounts thereof are 7 μm, and the total amount is 14 μm, there is no problem, and it can be further prevented. The gas generated by the gap leaks.

When the amount of elastic deformation is increased, gas leakage due to large bearing deformation amount or unpredictable bearing deformation amount can be prevented, and therefore, the selection range of the piston height clearance Δr becomes wider. For example, in Fig. 10, when R5 is used instead of R3, R5 can also obtain the same effect as R3.

As described above, the thin wall 38 disposed in the first embodiment has a proportional relationship with the pressure difference between the inside and the outside of the piston 28, and increases or decreases with the amount of elastic deformation, thereby greatly reducing the height of the piston generated from the operation of the compressor. The gap Δr leaks the re-expansion loss caused by the high-pressure gas to the cylinder compression chamber, and the effect of further improving the efficiency of the compressor can be achieved. At the same time, the selection range of the piston height gap Δr which can obtain the highest efficiency can be expanded, and the effect of improving production efficiency can be obtained.

When the compressor is stopped, the system is equivalent to the total pressure of the compressor. Therefore, in the first embodiment, the thin wall 38 at the time of starting the compressor is not elastically deformed. Therefore, the piston height gap Δr is the largest, and the frictional resistance caused by the sliding of the piston 28 is the smallest. Therefore, the compressor can be started more easily.

Fig. 11 shows a method in which the piston height gap Δr is previously reduced by the thin wall 38, that is, a method of performing the initial deformation of the quantitative D in advance in the direction in which the height of the piston is increased. According to this method, the operating pressure of the refrigerant is reduced, and even when the differential pressure between the discharge pressure Pd and the suction pressure Ps is small, the amount of elastic deformation can be made larger than the initial deformation amount of the quantitative D.

In the first embodiment, the thin walls constituting the upper and lower sides of the piston have the same shape; if the two thin wall shapes are changed, the amount of elastic deformation can be optimized. For example, it can be applied to a case where the deformation sizes of the main bearing 25 and the sub-bearing 26 are different.

As shown in FIG. 12, in order to increase the amount of elastic deformation of the thin wall 38, a method of providing the recess 39 in the inside of the circumferential groove 37 may be employed. The recess 39 is disposed in the circumferential groove 37 and abuts against the side of the thin wall 38.

As shown in Fig. 13, the thin wall on one side can be eliminated. In the design in which the one-side thin wall is eliminated, the amount of elastic deformation of the other side thin wall 38 is increased, and the intended action and effect can be exerted.

As shown in Fig. 14, in the fitting range of the eccentric shaft 17 and the inside of the piston 28, the width W of the circumferential groove 37 is enlarged, and the effect of reducing the weight of the piston 28 can be increased. Since the weight of the piston 28 is an unbalanced mass, the effect of reducing the vibration of the compressor can be obtained by reducing the weight.

Example 2

Referring to Fig. 15, the difference between the outer wall and the inner wall of the piston 28 is the design method of the thin wall 38 when the thickness t of the piston is thinned and the circumferential groove 37 is not deep enough.

In Fig. 15, the shaft diameter of the counter shaft 34b is designed to be smaller than the shaft diameter of the main shaft 34a of the crankshaft 27, and the width of the end portion of the piston toward the sub-bearing is larger than the width toward the end portion of the main bearing, thereby expanding the piston 28 Internally, a space portion formed between the eccentric shaft 17 and the sub-bearing 26. As a result, as in the first embodiment, since the entire length of the thin wall 38 on the lower side of the piston 28 can be increased, the depth of the circumferential groove 37 becomes shallow, and the problem that the amount of elastic deformation of the thin wall 38 is reduced can be solved.

The remaining parts are not shown in the first embodiment and will not be repeated.

Example 3

Referring to Figure 16, there is a feature of providing four oil holes 41 in the thin wall 38. One end of the oil hole 41 is bored in the end surface of the piston, and the other end of the oil hole is bored in the circumferential groove 37. The oil hole 41 can supply sufficient oil inside the piston 28 to the upper and lower sliding surfaces of the piston 28, and can further reduce gas leakage of these sliding surfaces, and has an effect of preventing friction loss and abrasion of the sliding surface.

In the present embodiment, by opening the oil hole 41 in the circumferential groove 37, there is a feature that oil can be directly supplied from the inside of the circumferential groove 37 to the sliding surface of the thin wall 38. The number of the oil holes 41 should be set reasonably, and can be appropriately increased or decreased.

As shown in Fig. 17, an annular groove 42 communicating with the oil hole 41 is added to the thin wall 38. As a result, the entire upper and lower sliding surfaces of the piston 28 can be lubricated, and the above effects can be further enhanced.

The remaining parts are not shown in the second embodiment and will not be repeated.

Example 4

Referring to Fig. 18, a lubricating film 32 is formed on the end faces of the main bearing 25 and the sub-bearing 26 on the upper and lower sliding surfaces of the piston 28, for example, using a material having excellent lubricity such as a molybdenum disulfide film or a ferric phosphate film. As a result, not only the friction coefficient when the thin wall 38 slides is improved, but also the convex portion of the lubricating film 32 is gradually worn off as the operation time advances, and the bearing deformation of the compression mechanism portion 21 is uniformized, and the flatness is improved. Therefore, gas leakage inside the piston caused by elastic deformation of the thin wall 38 can be further prevented. If the flatness is improved, the wear of the lubricating film 32 is stopped. The lubricating film 32 may be provided on both the main bearing 25 and the sub-bearing 26 at the same time, or may be separately provided on any one of them.

The remaining parts are not shown in the third embodiment and will not be repeated.

Example 5

Referring to Fig. 19, the technique disclosed in Embodiment 1 is applied to a two-cylinder rotary compressor (50). The techniques of Embodiments 2 to 4 can be applied relatively simply if necessary.

The twin-cylinder compression mechanism portion 51 of the two-cylinder rotary compressor 50 includes a first cylinder 52a and a second cylinder 52b, a middle plate 53 disposed therebetween, and a cylindrical first piston 54a and a second piston 54b respectively accommodated in the above two In the cylinder, two slides (not shown) that are in contact with the outer circumference of the pistons, a two-cylinder crankshaft 55 that drives the two pistons, a main bearing 56 that supports the two-cylinder crankshaft 55, and a sub-bearing 57. The twin-cylinder compression mechanism portion 51 is assembled by the above-described components and corresponding cylinder screws, wherein the outer circumference of the main bearing 56 is welded and fixed to the inner wall of the seal housing 60.

As in the first embodiment, the first piston 54a and the second piston 54b are respectively formed with thin walls 38 by circumferential grooves 37 provided in the vicinity of the upper and lower ends of the piston. In the operation of the compressor, the upper and lower piston height gaps formed between the main bearing 56 and the intermediate plate 53 of the upper and lower ends of the first piston 54a Δr and the upper and lower piston height gaps Δr formed between the sub-bearing 57 and the intermediate plate 53 of the upper and lower ends of the second piston 54b are respectively changed in accordance with the bearing deformation of the compression mechanism portion 51 according to the first embodiment; The elastic deformation of the thin wall 38 is minimized or optimized.

Therefore, the amount of gas leaking from the inside of the first piston 54a to the cylinder compression chamber can be reduced. Also, in the second piston 54b, leakage of gas can be reduced. As a result, as in the first embodiment, the efficiency of the compressor can be improved.

In order to improve the flatness, a lubricating film 32 may be provided on the end faces of the main bearing, the sub-bearing, and/or the intermediate plate 53 toward the piston side. In other words, it can be set on any of the above three components as needed, or it can be set in all of the above three components.

In the twin-cylinder rotary compressor, the number of components constituting the compression mechanism portion 51 is increased, and the overall height of the compression mechanism portion 51 is increased, so that the bearing deformation is higher than that of the single-cylinder rotary compressor of the first embodiment. Add a paragraph. Therefore, in the two-cylinder rotary compressor, a method of preventing leakage of high-pressure gas from the inner diameter of the piston is more effective.

In Fig. 19, the piston 54a and the piston 54b are each provided with a circumferential groove 37 at the upper and lower ends thereof, as described in Embodiment 1, in which one of the two pistons may be omitted as needed. Or change the shape design of the circumferential groove 37. Among the two cylinders, in the design of the height or the size of the cylinder compression chamber, various design changes such as changing the arrangement and design of the circumferential groove can be performed.

The remaining parts are described in the fourth embodiment and will not be repeated.

In summary, the present invention provides a piston and a main bearing by providing a circumferential groove near the upper and lower ends of the piston to form a thin wall, and following the pressure difference between the inside and the outside of the piston to elastically deform the thin wall. Loss of compressor efficiency caused by high pressure gas leaking from the sliding gap formed between the sub-bearing or the intermediate plate.

The technical solution disclosed by the present invention can be applied not only to a vertical rotary compressor of a crankshaft and a motor which are vertically disposed, but also to a horizontal rotary compressor of a crankshaft and a motor which are disposed laterally.

In summary, the technical solution disclosed in the present invention is easy to introduce into the industry and is easy to mass-produce; and the efficiency of the rotary compressor has a significant improvement.

Claims

A rotary compressor, in which a motor portion (22) and a compression mechanism portion (21) are disposed in a sealed casing (1), the compression mechanism portion includes more than one cylinder (23), and a cylinder compression chamber is provided in the cylinder (24) The piston (28) is received in the cylinder compression chamber, the tip end of the sliding plate is connected to the outer circumference of the piston, and the cylinder compression chamber is divided into a high pressure chamber (31a) and a low pressure chamber (31b), and the crank shaft (27) is connected to the piston and a driving piston, and a main bearing (25) and a sub-bearing (26) for supporting the crankshaft and mounted on the cylinder, the main bearing is connected to the main shaft (34a) of the crankshaft, and the sub-bearing is connected to the counter shaft (34b) of the crankshaft. , characterized in that at least one of the two end portions (36) of the piston is provided with a circumferential groove (37) at a position facing the end surface thereof, the opening of the circumferential groove facing the inside of the piston, and the outer side of the circumferential groove is provided with elasticity Thin wall (38); the gap formed between the two ends and the main bearing (25) and the sub-bearing (26) is adjusted by the pressure difference between the inside and the outside of the piston.
A rotary compressor according to claim 1, wherein said compression mechanism portion includes a first cylinder and a second cylinder, and an intermediate plate (53) is disposed between the first cylinder and the second cylinder, the first piston being disposed In the first cylinder, the second piston is disposed in the second cylinder, and the circumferential groove (37) is disposed on at least one of the four ends (36) of the first piston and the second piston; The difference in pressure between the inside and the outside of the piston and/or the second piston adjusts the gap formed between the ends of the first and second pistons with respect to any of the main bearing, the sub-bearing, and the intermediate plate.
A rotary compressor according to claim 1 or 2, wherein a shaft diameter of said main shaft (34a) of said crankshaft (27) is larger than a shaft diameter of said counter shaft (34b), and said piston faces an end portion of said sub-bearing. The width is greater than the width toward the end of the main bearing.
The rotary compressor according to claim 1 or 2, wherein the piston is provided with an oil hole (41), one end of the oil hole is opened at the end surface of the piston, and the other end of the oil hole is opened at the circumference Slot (37).
A rotary compressor according to claim 4, wherein an end surface of said piston is provided with an annular groove (42) which communicates with the oil hole (41).
A rotary compressor according to claim 1 or 2, wherein said circumferential groove (37) is provided with a recess (39) for increasing the amount of elastic deformation of the thin wall (38).
A rotary compressor according to claim 6, wherein said recess (39) is disposed in a side of the circumferential groove (37) against the thin wall (38).
A rotary compressor according to claim 1, wherein a lubricating film (32) is provided on an end surface of said main bearing (25) and/or sub-bearing (26) facing the piston.
A rotary compressor according to claim 2, wherein a lubricating film (32) is provided on an end surface of said main bearing (25), sub-bearing (26) and/or intermediate plate (53) facing the piston. .
A rotary compressor according to claim 1, wherein said rotary compressor is a horizontal rotary compressor or a vertical rotary compressor, said rotary compressor and condenser, expansion device and evaporator Form a refrigeration cycle.