KR101221311B1

KR101221311B1 - Swash plate type compressor

Info

Publication number: KR101221311B1
Application number: KR1020100089579A
Authority: KR
Inventors: 이정재; 이동주; 김민규; 윤덕빈; 손재식
Original assignee: 한라공조주식회사
Priority date: 2010-09-13
Filing date: 2010-09-13
Publication date: 2013-01-10
Also published as: KR20120027792A

Abstract

The present invention relates to a swash plate compressor in which a refrigerant suction passage having a diameter capable of maximizing compressor performance according to the maximum refrigerant discharge capacity of the compressor is formed. In the present invention, the suction passage 13 'is formed on the inner circumferential surfaces of the plurality of cylinder bores 12a, which are formed by coupling the front head 11 and the rear head 28 to the front and rear of the discharge chambers 11a and 28a. A passage through which the cylinder bore 12a is communicated by the suction passage 13 'to the rotating shaft 24 which is penetrated and rotatably installed through the front head 11 and the cylinder blocks 12 and 12'. 24 'is formed inside. In the cylinder bore 12a, a plurality of pistons 30 linearly reciprocating are included to compress the refrigerant. At this time, the diameter y of the suction passage 13 ′ is adjusted by a function of the maximum refrigerant discharge capacity x of the compressor 10. According to the present invention, the diameter of the suction passage is appropriately selected, so that the performance of the compressor can be improved, and the compressor can be easily designed.

Description

Swash plate type compressor {Swash plate type compressor}

The present invention relates to a swash plate compressor, and more particularly, to a swash plate compressor having a refrigerant suction passage having a diameter capable of maximizing compressor performance according to the maximum refrigerant discharge capacity of the compressor.

The compressor used in the automobile air conditioning system sucks the evaporated refrigerant from the evaporator and transfers it to the condenser at a high temperature and high pressure state which is easy to be liquefied.

In such a compressor, there is a reciprocating type in which compression is performed while reciprocating motion is actually performed for compressing the refrigerant, and a rotary type in which compression is performed while rotating. The reciprocating type includes a crank type for transmitting a driving force of a driving source to a plurality of pistons using a crank, a swash plate type for transferring a rotating shaft provided with a swash plate, and a wobble plate type using a wobble plate. Rotary types include rotary rotary axes with vane rotary vanes, scrolling with rotary scrolls and fixed scrolls.

1 is a sectional view of a general swash plate compressor, and FIG. 2 is a perspective view of a main part of a cylinder block according to the prior art.

As shown in the figure, the compressor 10 is coupled to a front cylinder block 12 and a rear cylinder block 12 'having a plurality of cylinder bores 12a, and coupled to the front of the front cylinder block 12. A front head 11 for forming the discharge chamber 11a and a rear head 28 for forming the discharge chamber 28a by being coupled to the rear of the rear cylinder block 12 'are included. They are arranged and combined in the order of the front head 11, the front cylinder block 12, the rear cylinder block 12 'and the rear head 28.

The front head 11 has a substantially cylindrical shape, and the discharge chamber 11a is formed therein. The discharge chambers 11a are open toward the front cylinder blocks 12, respectively. The discharge chamber 11a is formed over a substantially ring-shaped region. The discharge chamber 11a is formed to be selectively connected to each cylinder bore 12a of the front cylinder block 12 through a valve assembly 14 to be described below.

The front head 11 penetrates through the center of the shaft hole (O) is formed. The shaft shaft hole O is provided with a rotating shaft 24 to be described below.

The shaft support hole 13 is formed through the front cylinder block 12 and the rear cylinder block 12 '. The shaft 24 passes through the shaft support hole 13. The inner diameter of the shaft support hole 13 is designed to be in close contact with the outer surface of the rotary shaft 24.

The front cylinder block 12 and the rear cylinder block 12 'are coupled to the front head 11 and the rear head 28, respectively. A plurality of cylindrical cylinder bores 12a are formed in the front cylinder block 12 and the rear cylinder block 12 'in the direction of forming the shaft support hole 13 with the shaft support hole 13 as the center. .

The cylinder bore 12a and the shaft support hole 13 are connected to each other through a suction passage 13 ', respectively. The suction passages 13 ′ allow the refrigerant delivered through the inside of the rotating shaft 24 to be transferred to the cylinder bores 12a, respectively. As shown in FIG. 2, the suction passage 13 ′ has a cylinder bore adjacent to the front of the front and rear cylinder blocks 12, 12 ′ facing the front head 11 and the rear head 28, respectively. It is formed so as to penetrate toward the rotating shaft 24 in the inner peripheral surface of (12a).

Discharge passages (not shown) are formed in the front cylinder block 12 and the rear cylinder block 12 'so as to communicate with the discharge chambers 11a and 28a of the front head 11 and the rear head 28, respectively. The discharge passage serves as a passage for discharging the refrigerant compressed in the cylinder bore 12a to the outside.

The flow of the refrigerant is controlled between the discharge chamber 11a and the cylinder bore 12a between the front head 11 and the front cylinder block 12 and between the rear head 28 and the rear cylinder block 12 '. The valve assembly 14 is provided. That is, the valve assembly 14 controls the refrigerant flow from the cylinder bore 12a to the discharge chamber 11a.

The valve assembly 14 includes a valve plate 15 having a discharge hole 15 ′ and a discharge lead 17 coupled to one surface of the valve plate 15 facing the discharge chambers 11a and 28a. Include. The valve plate 15 has a substantially disc shape and has a discharge hole 15 'through which refrigerant is discharged at a position corresponding to each cylinder bore 12a. In addition, the discharge lead 17 is a material capable of elastic deformation and elastically deforms according to the internal pressure of the cylinder bore 12a to open and close the discharge hole 15 '.

A retainer 18 is provided on one surface of the valve plate 15 facing the front head 11. The retainer 18 is formed in a substantially plate shape, and is bent at a predetermined angle toward the discharge chamber 11a. The retainer 18 is excessively elastic toward the inside of the discharge chamber 11a of the front head 11 and the discharge chamber 28a of the rear head 28 by the discharge pressure of the refrigerant. This is to prevent deformation.

The front cylinder block 12 and the rear cylinder block 12 ′ are formed with recessed portions formed on surfaces coupled to each other to form the swash plate chamber 23. In the swash plate chamber 23, the swash plate 26 provided on the rotating shaft 24 is rotatably positioned.

A rotating shaft 24 is installed through the center of the front head 11, the front cylinder block 12 and the rear cylinder block 12 '. The flow passage 24 ′ through which the refrigerant flows is formed in the rotation shaft 24. The flow passage 24 ′ is elongated in the longitudinal direction of the rotation shaft 24 inside the rotation shaft 24. An inlet 24a and an outlet 24b are formed on the outer surface of the rotation shaft 24. The inlet 24a connects the swash plate chamber 23 and the flow path 24 ', and the outlet 24b is the suction passage 13' of the front cylinder block 12 and the rear cylinder block 12 '. It is formed at a position that can be connected to). The position of the outlet 24b should be formed in accordance with the compression order of the refrigerant proceeding in each cylinder bore 12a.

The shaft seal 25 is inserted into one side of the rotation shaft 24 to be in close contact with the inner surface of the shaft through hole (O) of the front head (11). The shaft seal 25 serves to prevent the refrigerant from leaking between the rotating shaft 24 and the shaft through hole (O). The shaft seal 25 is formed of a rubber material capable of elastic deformation.

An approximately disk-shaped swash plate 26 is inclined with respect to the extending direction of the rotating shaft 24 on the rotating shaft 24. A plurality of hemispherical shoes 27 surrounding the edge of the swash plate 26 are provided. Therefore, when the swash plate 26 having a predetermined inclination rotates and its edge portion passes the shoe 27, the piston 13 connected to the shoe 27 by the inclination of the swash plate 26 causes the cylinder bore. The refrigerant is compressed while linearly reciprocating in the interior of 12a.

On the other hand, a piston 30 is installed in the cylinder bore 12a so as to reciprocate linearly. The piston 30 is positioned in the cylinder bore 12a of the front cylinder block 12 and the rear cylinder block 12 'in an approximately cylindrical shape corresponding to the interior of the cylinder bore 12a. That is, the piston 30 serves to compress the refrigerant in the cylinder bore 12a. The piston 30 is the middle portion thereof is coupled to the shoe 27, the linear reciprocating motion according to the rotation of the swash plate 26.

The rear head 28 is mounted on one surface of the rear cylinder block 12 '. The rear head 28 is formed with a discharge chamber 28a. The discharge chamber 28a is formed over a substantially ring-shaped area. The discharge chambers 28a are open toward the rear cylinder blocks 12 ', respectively. The discharge chamber 28a is selectively connected to the cylinder bores 12a formed in the rear cylinder block 12 'through the valve plate 15.

The pulley 40 is rotatably installed at one side of the front head 11. The pulley 40 is formed in a substantially cylindrical shape. The pulley 40 is rotated by receiving the driving force of the engine through a belt (not shown).

The field coil 41 is built in the pulley 40. The field coil 41 generates suction magnetic flux when the power is applied to allow the disk 46 to be described below to closely adhere to the friction surface 40 ′ of the pulley 40.

Meanwhile, a hub 43 is installed at one end of the rotation shaft 24, and a damper 44 is installed at the hub 43. The damper 44 absorbs the shock generated during power transmission between the rotary shaft 24 and the pulley 40. The damper 44 is provided to move the disk 46 in a position facing the friction surface 40 ′ of the pulley 40.

The operation of the compressor having such a structure will be described. When the driving force of the engine is transmitted to the pulley 40 through the belt, the pulley 40 is rotated. However, when power is not applied to the field coil 41, the disk 46 is not in close contact with the friction surface 40 ′ of the pulley 40, so that the rotation shaft 24 does not rotate.

In this state, when the necessity of operation of the air conditioning system occurs and the compressor needs to be driven, the control system of the user or the vehicle provides a signal for the operation of the air conditioning system. When the operation of the air conditioning system is started and the refrigerant needs to be compressed, the field coil 41 generates suction flux while power is applied to the field coil 41.

When power is applied to the field coil 41, the disk 46 is in close contact with the friction surface 40 ′ of the pulley 40 by the suction magnetic flux of the field coil 41. Accordingly, the rotational force of the pulley 40 is transmitted to the rotation shaft 24 through the disk 46, the damper 44 and the hub 43.

When the rotational force of the pulley 40 is transmitted to the rotation shaft 24 as described above, the rotation shaft 24 rotates to linearly reciprocate the piston to compress the refrigerant.

At this time, as the rotary shaft 24 rotates, the flow passage 24 'inside the rotary shaft 24 is connected to the cylinder bore 12a through the outlet 24b and the suction passage 13'. The connection of the flow path 24 'and the cylinder bore 12a allows the refrigerant sucked into the swash plate chamber 23 to be transferred to the cylinder bore 12a. For reference, the refrigerant is sucked into the cylinder bore 12a when the piston 30 is located at the bottom dead center of the corresponding cylinder bore 12a. As such, when the refrigerant is delivered to the cylinder bore 12a, the piston 30 of the corresponding cylinder bore 12a moves in the direction of the valve plate 15, and the refrigerant is compressed.

Thus, when the refrigerant is compressed in the cylinder bore 12a, the pressure inside the cylinder bore 12a becomes relatively high, and the refrigerant is discharged to the discharge chambers 11a and 28a. The refrigerant discharged into the discharge chambers 11a and 28a is transferred to a condenser (not shown) through an external discharge port.

The refrigerant delivered to the condenser through the discharge port is delivered to the compressor again through a condenser (not shown), an expansion valve (not shown), and an evaporator (not shown). In the compressor the refrigerant is compressed by repeating the process described above.

However, the conventional compressor as described above has the following problems.

The suction passage 13 'is formed on the inner circumferential surface of the cylinder bore 12a adjacent to front surfaces of the front and rear cylinder blocks 12, 12'. At this time, if the diameter of the suction passage 13 'is small, the flow rate of the refrigerant sucked into the cylinder bore 12a is insufficient, and the performance of the compressor is reduced. If the diameter of the suction passage 13' is large, the suction passage ( 13 ') As the dead volume corresponding to the volume inside increases, the compression performance of the compressor decreases. That is, if the diameter of the suction passage 13 'is not properly adjusted, there is a problem that the performance of the compressor 10 is lowered.

Accordingly, an object of the present invention is to solve the problems of the prior art as described above, and to provide a swash plate compressor having a suction passage having a diameter capable of improving the performance of the compressor.

Another object of the present invention is to provide a swash plate compressor having a suction passage of an appropriate diameter according to the discharge capacity of the compressor.

According to a feature of the present invention for achieving the above object, the swash plate type compressor according to the present invention includes a front head and a rear head are respectively formed with a discharge chamber through which the refrigerant is discharged; A cylinder block in which the front head and the rear head are respectively coupled to the front and the rear, and a plurality of cylinder bores are formed, and an inlet passage having an opening is formed in the inner circumferential surface of the cylinder bore; A rotating shaft rotatably installed through the front head and the cylinder block, the swash plate being rotatable, and a flow passage communicating with the cylinder bore by the suction passage formed therein; And a plurality of pistons linearly reciprocating in the cylinder bore according to the rotational motion of the swash plate, wherein the diameter y of the suction passage is the maximum refrigerant discharge capacity x of the compressor. Equation for

(only,

And a value within a range calculated by x≥110.

Here, when the maximum refrigerant discharge capacity of the compressor is 110cc, the diameter y of the suction passage may have a value within the range of 4.55mm to 4.85mm.

When the maximum refrigerant discharge capacity of the compressor is 130 cc, the diameter y of the suction passage may have a value within the range of 5.35 mm to 5.65 mm.

Further, when the maximum refrigerant discharge capacity of the compressor is 160cc, the diameter y of the suction passage may have a value within the range of 5.7mm to 6.0mm.

The swash plate compressor according to the present invention has the following effects.

That is, there is an advantage that the performance of the compressor can be improved by appropriately selecting the diameter of the suction passage.

Furthermore, according to the swash plate compressor according to the present invention, since the diameter of the suction passage is calculated according to the discharge capacity of the compressor, there is an advantage that the compressor design is easy.

1 is a cross-sectional view showing the configuration of a typical swash plate compressor.
Figure 2 is a perspective view showing the main configuration of the cylinder block of a typical swash plate compressor.
3 is a graph showing the change in compressor performance according to the suction passage diameter of the compressor having a discharge capacity of 110cc.
4 is a graph showing changes in compressor performance according to the suction passage diameter of the compressor having a discharge capacity of 130 cc.
5 is a graph showing changes in compressor performance according to the suction passage diameter of a compressor having a discharge capacity of 160 cc.

Hereinafter, the configuration of a preferred embodiment of the swash plate compressor according to the present invention will be described in detail with reference to the drawings. Since the configuration of the swash plate compressor according to the embodiment of the present invention is the same as described with reference to FIGS. 1 and 2 in the background art, a detailed description thereof will be omitted and the same components will be indicated by using the same reference numerals.

In the present invention, in order to calculate a preferable value of the diameter of the suction passage 13 'formed in the front and rear cylinder blocks 12, 12' of the compressor 10, a compressor having different maximum refrigerant discharge capacities. Calculate the appropriate suction passage 13 'diameter of (10).

3 is a graph showing a change in compressor performance according to the suction passage diameter of the compressor having a discharge capacity of 110cc, Figure 4 is a graph showing a change in compressor performance according to the suction passage diameter of a compressor having a discharge capacity of 130cc, Figure 5 is a discharge This is a graph showing changes in compressor performance according to the suction passage diameter of a compressor with a capacity of 160cc.

Here, the maximum refrigerant discharge capacity of the compressor 10 refers to the amount of refrigerant that can be discharged from the compressor 10 while the swash plate 26 provided in the compressor 10 is rotated one time. In general, when the maximum refrigerant discharge capacity of the compressor 10 is increased, the overall volume of the compressor 10 also increases.

First, the graph shown in the upper part of FIG. 3 shows that the diameter of the suction passage 13 ′ is changed while changing the revolutions per minute of the compressor 10 to 800, 1800, and 3000 RPM in the compressor 10 having a maximum refrigerant discharge capacity of 110 cc. Is a graph showing a change in the cooling capacity (Capacity, Kcal / h) of the compressor 10 according to the change.

First, when the rotational speed per minute of the compressor 10 is 800 RPM, as the diameter of the suction passage 13 ′ increases, the cooling capacity of the compressor 10 may be slightly reduced. In addition, when the rotation speed per minute of the compressor 10 is 1800 RPM, the cooling capacity of the compressor 10 decreases as the diameter of the suction passage 13 ′ increases. However, when the rotational speed per minute of the compressor 10 is 3000 RPM, the cooling performance of the compressor 10 also increases rapidly as the diameter of the suction passage 13 'is increased.

That is, when the diameter of the suction passage 13 'is about 4.4 mm, the compressor cooling performance when the compressor 10 is driven at high speed is lower than when the suction passage 13' is different, and the diameter of the suction passage 13 'is about 5.0. In the case of mm or more, the cooling performance at the time of high speed operation of the compressor 10 is good, but it can be seen that the cooling performance at the time of low speed operation is lower than other cases. On the other hand, in the section where the diameter of the suction passage 13 'is about 4.7 mm, both the high speed operation and the low speed operation of the compressor 10 show good cooling performance.

On the other hand, the graph shown at the bottom of Figure 3, the compressor 10 while changing the number of revolutions per minute of the compressor 10 to 800, 1800, 3000 RPM in the compressor 10 having a maximum refrigerant discharge capacity of 110cc as above, It shows the result of observing the change of COP (Coefficient of Performance). Here, the performance coefficient is a numerical value representing the refrigeration effect by the compressor 10 on the work supplied to the compressor 10.

When the rotation speed per minute of the compressor 10 is 800 RPM, as the diameter of the suction passage 13 'increases, the coefficient of performance of the compressor 10 also gradually increases, and the diameter of the suction passage 13' is 4.7 mm. We can see that we have a maximum performance factor and then gradually decrease again. In addition, when the rotation speed per minute of the compressor 10 is 1800 RPM, as the diameter of the suction passage 13 ′ increases, the coefficient of performance of the compressor 10 decreases. When the number of revolutions per minute of the compressor 10 is 3000 RPM, as the diameter of the suction passage 13 'is increased, the coefficient of performance of the compressor 10 increases.

In other words, when the diameter of the suction passage 13 'is approximately 4.7 mm, the compressor 10 exhibits good performance coefficients when the compressor 10 is operated at high speed and at low speed, but the diameter of the suction passage 13' is small. In the case of high speed operation, the performance is poor, and in the case of the diameter of the suction passage 13 'being large, the coefficient of performance in the low speed operation decreases.

Therefore, based on the experimental results shown in the two graphs shown in FIG. 3, if the proper diameter of the suction passage 13 ′ of the compressor 10 having a maximum refrigerant discharge capacity of 110 cc is derived, the diameter becomes about 4.7 mm.

On the other hand, the graph shown in the upper portion of Figure 4 is the compressor 10 of the maximum refrigerant discharge capacity 130cc of the suction passage 13 'while changing the revolutions per minute of the compressor 10 to 1000, 1800, 3000RPM A graph showing a result of observing a change in the cooling capacity of the compressor 10 according to the change in diameter.

In the low-speed operation in which the rotational speed of the compressor 10 is 1000 RPM, it can be seen that there is no significant difference in the cooling capacity of the compressor 10 even if the diameter of the suction passage 13 ′ is changed. In addition, even when the rotational speed per minute of the compressor 10 is 1800 RPM, the change in the cooling capacity of the compressor 10 due to the change in the diameter of the suction passage 13 'is insignificant. However, when the number of revolutions per minute of the compressor 10 is 3000 RPM, the cooling performance of the compressor 10 also increases as the diameter of the suction passage 13 'increases.

According to this, when the diameter of the suction passage 13 ′ is about 5.3 mm, the cooling capacity of the compressor 10 is low in both the low speed operation and the high speed operation of the compressor 10, and the suction path 13 When the diameter of ') is approximately 5.7 mm, the cooling capacity of the compressor 10 is relatively high when the compressor 10 is operated at high speed.

On the other hand, the graph shown at the bottom of Figure 4, the compressor 10 while changing the number of revolutions per minute of the compressor 10 to 1000, 1800, 3000 RPM in the compressor 10 having a maximum refrigerant discharge capacity of 130cc as above, It shows the result of observing the change of COP (Coefficient of Performance).

When the rotation speed per minute of the compressor 10 is 1000 RPM, the performance coefficient of the compressor 10 gradually increases as the diameter of the suction passage 13 ′ increases, and the diameter of the suction passage 13 ′ is approximately 5.5. It can be seen that with mm the maximum performance factor is gradually reduced again. This is likewise shown when the revolutions per minute of the compressor 10 are 1800 RPM and 3000 RPM.

That is, when the diameter of the suction passage 13 'is about 5.5 mm, the best performance coefficient is shown regardless of the operating speed of the compressor 10.

Therefore, based on the experimental results shown in the two graphs shown in FIG. 4, an appropriate diameter of the suction passage 13 ′ of the compressor 10 having a maximum refrigerant discharge capacity of 130 cc is about 5.5 mm.

Furthermore, in FIG. 5, the changes in the cooling capacity and the coefficient of performance of the compressor 10 according to the revolutions per minute of the compressor 10 having the maximum refrigerant discharge capacity of 160 cc and the diameter of the suction passage 13 ′ are first examined. In the upper portion, the low-speed operation at 1000 RPM per minute of the compressor 10, even if the diameter of the suction passage 13 'is changed, there is no significant difference in the cooling capacity of the compressor 10, whereas the compressor 10 It can be seen that the cooling capacity of the compressor 10 decreases as the diameter of the suction passage 13 ′ increases when the number of revolutions per minute is 1800 RPM. In addition, when the number of revolutions per minute of the compressor 10 is 3000 RPM, as the diameter of the suction passage 13 ′ increases, the cooling capability of the compressor 10 may be improved.

According to this, when the diameter of the suction passage 13 'is approximately 5.5mm, the cooling capacity of the compressor 10 is relatively low when the compressor 10 is driven at high speed, and the diameter of the suction passage 13' is relatively low. In the case of approximately 6.1 mm, the cooling capacity of the compressor 10 is relatively low when the compressor 10 is operated at low speed.

As shown in the lower part of FIG. 5, in the compressor 10 having a maximum refrigerant discharge capacity of 160 cc, the performance coefficient of the compressor 10 is changed while changing the revolutions per minute of the compressor 10 to 1000, 1800, and 3000 RPM. Observing the change in COP (Coefficient of Performance), when the rotation speed per minute of the compressor 10 is 1000 RPM, as the diameter of the suction passage 13 'increases, the performance coefficient of the compressor 10 also gradually increases, It can be seen that when the diameter of the suction passage 13 'is approximately 5.85 mm, the suction passage 13' gradually decreases again with the maximum coefficient of performance. This is the same even when the revolutions per minute of the compressor 10 is 1800 RPM.

When the number of revolutions per minute of the compressor 10 is 3000 RPM, as the diameter of the suction passage 13 'increases, the coefficient of performance of the compressor 10 continuously increases, but considering both high speed operation and low speed operation. When the suction passage 13 'has a diameter of approximately 5.85mm it can be seen that the performance of the compressor 10 is the best.

Therefore, it is preferable that the suction passage 13 'of the compressor 10 having the maximum refrigerant discharge capacity of 130 cc is about 5.85 mm based on the experimental results shown in the two graphs shown in FIG.

As described above, while adjusting the maximum refrigerant discharge capacity of the compressor 10 in several steps, the diameter of the suction passage 13 ′ corresponding thereto was calculated. In the following, a general equation for calculating the diameter of the suction passage corresponding to the maximum refrigerant discharge capacity of the compressor is derived.

When the suction passage 13 'is large in diameter, the cross-sectional area of the suction passage 13' is widened to increase the dead volume of the compressor 10, and when the diameter of the suction passage 13 'is small, the suction passage 13' is smaller. Since the cross-sectional area of 13 'is narrowed to increase the resistance of the refrigerant sucked into the cylinder bore 12a, it is difficult to inhale the refrigerant and more work is required to suck a sufficient amount of refrigerant into the cylinder bore 12a. By making it large, the efficiency of the said compressor 10 is reduced.

At this time, when a diameter of said y, the suction passage (13, the suction passage 13, the cross-sectional area) is proportional to y ^2. The dead volume generated by the suction passage 13 ′ is also proportional to y ² , and the resistance of the refrigerant in the suction passage 13 ′ is inversely proportional to y ² .

As described above, when the maximum refrigerant discharge capacity of the compressor 10 is increased, the volume of the compressor 10 increases together. Accordingly, as the maximum refrigerant discharge capacity of the compressor 10 increases, the volume inside the compressor 10 increases, so that the size of the dead volume allowed based on the same efficiency also increases.

However, as the maximum refrigerant discharge capacity of the compressor 10 increases, not only the volume inside the compressor 10 increases but also the amount of refrigerant discharged increases, so that the amount of refrigerant to be sucked into the cylinder bore 12a also increases. . Therefore, as the maximum refrigerant discharge capacity of the compressor 10 increases, the amount of refrigerant sucked into the suction passage 13 'must increase, and for this purpose, the resistance of the refrigerant sucked into the suction passage 13' must be reduced. do.

That is, as the maximum refrigerant discharge capacity of the compressor 10 increases, the amount of refrigerant sucked into the cylinder bore 12a is smaller than the loss of the compressor efficiency due to the increase in the dead volume of the compressor 10, or the cylinder. Since the compressor efficiency deterioration caused by the amount of work applied to suck the refrigerant into the bore 12a is so large that the compressor 10 suffers some damage to the dead volume as the maximum refrigerant discharge capacity increases. Even if it is more important to ensure the area of the sufficient suction passage (13 ').

In other words, when the maximum refrigerant discharge capacity of the compressor 10 is x, as x increases, it is inversely proportional to the resistance of the suction passage 13 'and y ² proportional to the dead volume increases. . That is, x is proportional to y ² .

Therefore, x is expressed as a quadratic function for y. That is, when x = ay ² + by + c, the optimum diameter y of the suction passage 13 'when the maximum refrigerant discharge capacity x of the compressor 10 is 110, 130, and 160 cc is 4.7, Since we know that it is 5.5 and 5.85, and substituting them into the above quadratic functions to obtain the constants a, b and c, a is 52.795, b is -513.484 and c is calculated as 1357.175.

Thus x = 52.795y ² -513.484y + 1357.175. Therefore, if this is expressed as a function representing the diameter y of the suction passage 13 'of the compressor 10 according to the change of the maximum refrigerant discharge capacity x of the compressor 10, x = 52.795y ² -513.484y + It is the inverse of 1357.175.

That is, it becomes the following relationship.

Where x is the maximum refrigerant discharge capacity of the compressor 10, which is greater than at least 108.638 cc, and y is the optimum diameter value of the suction passage 13 'with respect to the maximum refrigerant discharge capacity of the compressor 10. The unit is mm.

That is, the diameter of the suction passage 13 'of the compressor 10 with respect to the maximum refrigerant discharge capacity x of the compressor 10 is calculated as described above. According to this, two different y values may be calculated with respect to one x value, but selecting one of the two values is the suction passage 13 ′ of the compressor 10 having the corresponding maximum refrigerant discharge capacity. It can be determined by calculating and comparing the efficiency with each of the two diameters.

The scope of the present invention is not limited to the embodiments described above, but may be defined by the scope of the claims, and those skilled in the art may make various modifications and alterations within the scope of the claims It is self-evident.

10: Compressor 11: Fronthead
11a: discharge chamber 12, 12 ': front and rear cylinder blocks
12a: cylinder bore 13: shaft support
13 ': suction passage 14: valve assembly
15: valve plate 15 ': discharge hole
23: swash chamber 24: axis of rotation
24 ': Euro 26: Saphan
27: shoe 28: rear head
28a: discharge chamber 30: piston
40: pulley 43: hub

Claims

A front head 11 and a rear head 28 each having discharge chambers 11a and 28a through which refrigerant is discharged;
The front head 11 and the rear head 28 are respectively coupled to the front and rear, a plurality of cylinder bores (12a) are formed, the suction passage 13 'having an opening in the inner circumferential surface of the cylinder bore (12a) Cylinder blocks 12 and 12 'are formed;
It is installed rotatably through the front head 11 and the cylinder block (12, 12 '), the swash plate 26 is rotatably installed, the cylinder bore (12a) by the suction passage (13') A rotating shaft 24 having a flow passage 24 ′ communicating therewith; And
In the swash plate-type compressor 10, characterized in that it comprises a plurality of pistons (30) for linear reciprocating motion in the cylinder bore (12a) in accordance with the rotational movement of the swash plate (26),
The diameter y of the suction passage 13 'is expressed by the equation for the maximum refrigerant discharge capacity x of the compressor 10.

(only,

And a value within a range calculated by x≥110).

The method of claim 1,
When the maximum refrigerant discharge capacity of the compressor 10 is 110cc,
The diameter y of the suction passage (13 ') is a swash plate compressor, characterized in that it has a value in the range of 4.55mm to 4.85mm.

The method of claim 1,
When the maximum refrigerant discharge capacity of the compressor 10 is 130cc,
The diameter y of the suction passage (13 ') is a swash plate compressor, characterized in that it has a value in the range of 5.35mm to 5.65mm.

The method of claim 1,
When the maximum refrigerant discharge capacity of the compressor 10 is 160cc,
The swash plate compressor, characterized in that the diameter y of the suction passage (13 ') has a value in the range of 5.7mm to 6.0mm.