JPH1026425A - Refrigerant compressor driving at variable speed and refrigeration cycle device provided with the same refrigerant compressor - Google Patents

Abstract

PROBLEM TO BE SOLVED: To minimize the ratio of leakage loss to motor loss against theoretical compression work of a refrigerant compressor and enhance the compression performance and execute fine control near room temperature by increasing the minimum value of a specified operation frequency zone during steady state operation except for starting time. SOLUTION: Except for starting time, a steady state operation frequency is transferred to a higher frequency side on the whole where a refrigerant compressor 100 is adapted to reduce its stroke volume so as to comply with a double-increased speed operation, thereby reducing the rate of leakage loss to theoretical compression work and enhancing dramatically the performance of the compressor 100 during a minimum frequency operation. The drive torque of a motor required for the compression of working refrigerant during the same refrigerating capacity is favorably half since the frequency is boubled. The operation at a low frequency less than a commercial power frequency, at which breakdown torque, is dramatically reduced, is not executed with in a steady state operation range. Therefore, motor loss is not generated and iron loss is reduced so that motor efficiency is enhanced and the performance of the compressor and period energy consumption efficiency of a refrigeration cycle device is improved, which makes it possible to execute sophisticated control near room temperature.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a refrigerant compressor which is used in a refrigeration cycle and performs variable speed driving by changing the frequency of a driving power supply by an inverter or the like. The present invention relates to a refrigeration cycle device that controls the operating frequency of a refrigerant compressor and performs variable speed driving of the refrigerant compressor.

[0002]

2. Description of the Related Art Conventionally, the minimum operating frequency of a refrigerant compressor used in a refrigeration cycle apparatus in which a refrigerant compressor is driven at a variable speed by an inverter or the like is disclosed in Japanese Patent Application Laid-Open No. 7-294030. Is about 28 Hz, which is the frequency of the commercial power supply (60 Hz or 50 Hz in Japan).
Hz).

As a refrigerant compressor used in a refrigeration cycle apparatus that drives a refrigerant compressor at a variable speed, a rotary compressor as shown in FIG. 7 or a scroll compressor as shown in FIG. Machine is common.

FIG. 7 is a longitudinal sectional view showing an example of a rotary compressor, and shows a schematic structure. FIG. 8 is a cross-sectional view of the compression element of FIG. A compression element 2 and an electric motor 3 corresponding to a variable speed for driving the compression element 2 at a variable speed via a crankshaft 4 are built in a closed container 1. Reference numeral 5 denotes a suction pipe for sucking the working refrigerant into the compression element 2, and is connected to a cylinder 6 forming a compression chamber 11. As shown in FIG. 8, the low pressure chamber 11a of the compression chamber 11, which is the internal space of the suction pipe 5 and the cylinder 6, is provided in the cylinder 6.
The cylinder 6 is provided with a suction hole 7 communicating therewith.
A ring piston 8 is provided so as to roll along the inner surface of the cylinder 6. Reference numeral 9 denotes a vane which is always pressed against the outer peripheral side surface of the rolling piston 8 and is attached to a vane groove 10 formed in the cylinder 6 so as to be able to reciprocate in the radial direction. It is separated into a low pressure chamber 11a. Further, a discharge hole 12 is formed in a cylinder portion located near the vane 9 opposite to the suction hole 7 and communicates with the high-pressure chamber 11b and discharges the compressed working refrigerant into the closed container 1. The hole 12 is opened and closed by a discharge valve 13.

In the rotary compressor constructed as above, the working refrigerant sucked into the low-pressure chamber 11a from the suction pipe 5 through the suction hole 7 is driven by the crankshaft 4.
The ring piston 8 is gradually compressed to a predetermined pressure by the eccentric rotation of the ring piston 8, discharged from the high-pressure chamber 11 b through the discharge hole 12, opens the discharge valve 13, and is discharged into the sealed container 1.

FIG. 9 shows a working refrigerant in which one spiral scroll tooth is stationary and the other spiral scroll tooth is revolved around the center of the stationary scroll tooth without rotating. 1 shows the basic components of a revolving type scroll compressor that compresses the pressure and the principle of compression. 2 in the figure
Reference numeral 1 denotes a fixed scroll fixedly fixed to a space and has spiral scroll teeth 21a. Reference numeral 22 denotes a revolving scroll having a winding direction opposite to that of the fixed scroll teeth 21a, and 180 degrees with respect to the fixed scroll teeth 21a.
゜ It has spiral scroll teeth 22a that are combined in a phase shifted state. These spiral scroll teeth 2
The shapes of 1a and 22a are a combination of involute curves and arcs. 23 is a suction chamber, 24 is a discharge hole, and 25 is a compression chamber.

Next, the operation of FIG. 9 will be described. The orbiting scroll 22 performs a rotation motion, that is, a revolving motion around the center of the fixed scroll teeth 21a without changing its posture, that is, without rotating, and FIG.
0 °, 90 °, 180 shown in (b), (c), and (d)
It moves sequentially like a working position of {270}. In the state of 0 ° shown in FIG. 9 (a), the closing of the working refrigerant in the suction chamber 23 is completed, and the fixed scroll teeth 21a and the swirl scroll are closed.
A crescent-shaped compression chamber 25 is formed between the teeth 22a. Then, with the revolving motion of the orbiting scroll 22, the compression chamber 2
Numeral 5 sequentially reduces the volume, compresses the working refrigerant to the vicinity of the center of the fixed scroll 21, and discharges it from the discharge hole 24.

Next, a specific configuration and operation of the scroll compressor will be described. FIG. 10 is a longitudinal sectional view showing an example of a revolving type scroll compressor.
Is a fixed scroll provided with spiral scroll teeth 21a on a base plate 21b, which is fixed to the main frame 26. 22
Is a revolving scroll provided with spiral scroll teeth 22a combined with the fixed scroll 21a on the base plate 22b, and is provided on the end face of the base plate 22b opposite to the side on which the scroll teeth 22a are provided. A thrust surface 22c which receives a thrust force due to a compressive load acting on the orbiting scroll 22, and a oscillating bearing 2 which supports a radial load such as a compressive load acting on the orbiting scroll 22 and a centrifugal force generated by revolving motion.
2d is formed. The main frame 26 is fixed to a closed container 27, and a thrust bearing 26a is formed on the main frame 26 to support the thrust surface 22c of the revolving scroll 22 in the axial direction so as to be slidable. I have. Reference numeral 28 denotes a variable speed motor that drives the main shaft 29 to rotate at a variable speed, and includes a stator 28a and a rotor 28b. The motor rotor 28b is shrink-fitted and fixed to the main shaft 29, and balances the entire rotating system against the centrifugal force generated by the revolving motion of the orbiting scroll 22.
0b is attached to the upper part and the lower part. The main shaft 29 is rotationally driven by an electric motor 28, and is supported in a radial direction on both sides of the electric motor 28 by a main bearing 26b formed on the main frame 26 and a sub-bearing 31a formed on the sub-frame 31. ing. Numeral 32 denotes an Oldham ring which reciprocates linearly relative to both the revolving scroll 22 and the main frame 26. Eccentric shaft 29 at the upper end of main shaft 29
The eccentric shaft 29a is driven by the rotation of the main shaft 29 to rotate the eccentric shaft 29a through the swing bearing 22d.
2, the rotating scroll 22 is rotated by the reciprocating linear motion of the Oldham ring 32, so that the revolving scroll 22 revolves with respect to the fixed scroll 21 to compress the working refrigerant as described above. You do it. In addition, sealed container 27
Is formed with a lubricating oil reservoir 33, and a lubricating oil reservoir 3 is provided by a positive displacement pump 34 attached to the lower end of the main shaft 29.
The lubricating oil of No. 3 is supplied to bearings 22d, 26a, 26b, 31a and other sliding parts through oil passages 29b formed in the main shaft 29.

[0009]

SUMMARY OF THE INVENTION Refrigerant compressors such as the rotary compressor and the scroll compressor described above are disclosed in Japanese Patent Laid-Open No. 7-2940 in a steady operation range except at the time of starting.
As shown in Japanese Patent Publication No. 30, when the compressor is operated at a low frequency lower than the commercial power frequency such as 28 Hz, the performance of the refrigerant compressor is significantly reduced as compared with the case where the refrigerant compressor is operated at the commercial power frequency or a high frequency higher than the commercial power frequency. There is a problem.

The first reason is that the ratio of leakage loss to the theoretical compression work increases during low-speed operation. This is because the amount of internal leakage per unit time during which the working refrigerant leaks during compression or after completion of compression through various clearances to the compression chamber on the lower pressure side is independent of the rotation speed (that is, the operating frequency). It is. Therefore, the time required for one rotation increases as the operation speed decreases, so that the amount of leakage during one rotation increases, and the ratio of leakage loss to the theoretical compression work increases under the same pressure condition, and the coefficient of performance decreases. The various gaps where leakage occurs include, in a rotary compressor, the side surfaces of the vanes and the vane grooves as paths through which the working refrigerant at the discharge pressure once discharged into the closed vessel leaks into both the high-pressure chamber and the low-pressure chamber. There is an axial clearance at the upper end surface of the rolling piston that has been left, and as the leakage path from the high-pressure chamber to the low-pressure chamber, the axial direction of the upper end surface of the vane that has been left in the radial direction between the outer peripheral surface of the rolling piston and the inner peripheral surface of the cylinder. There is a gap. In a scroll compressor, as a leakage path to an adjacent compression chamber having a pressure lower than its own compression chamber pressure, both of the spiral teeth which are left axially between the tooth tip surface of the spiral tooth and the end face of the base plate of the other scroll are used. There is a radial clearance between the sides. Further, in the scroll compressor, the working refrigerant in the compression chamber formed at the outermost periphery may leak to the outside of the outermost compression chamber from the above-mentioned leakage path at an early stage of the compression stroke. It decreases significantly.

The second reason is that during low-speed operation, the ratio of motor loss to the theoretical compression work is high. Usually, a three-phase squirrel-cage induction motor is used as the motor of the refrigerant compressor that performs variable speed driving by changing the frequency of the driving power supply with an inverter or the like. In a three-phase squirrel-cage induction motor, if the magnetic flux density of the magnetic steel sheet forming the stator or rotor or the magnetic flux density in the gap between the stator and rotor is to be constant regardless of frequency, the relationship between voltage and frequency is In this case, the stall torque becomes smaller as the frequency becomes lower. In this case, there is a point where the stall torque rapidly decreases particularly at a low frequency lower than the commercial power frequency. Therefore, when an attempt is made to operate under the same pressure condition as the frequency equal to or higher than the commercial power supply frequency, the operation torque approaches the stop torque as the frequency is lower (lower speed), and the current value increases in such a case. The loss increases (copper loss is proportional to the square of the current value), and the motor efficiency decreases. That is, the ratio of the motor loss to the theoretical compression work increases. Therefore, especially at a low frequency lower than the commercial power supply frequency at which the stall torque is reduced, a method is employed in which the voltage is higher than a voltage value determined by the proportional relationship in order to increase the stall torque, whereby the motor efficiency is higher than the above. However, even in this case, increasing the voltage increases the magnetic flux density of the magnetic steel sheets constituting the stator and the rotor, and the magnetic flux density in the clearance between the stator and the rotor. (Proportional to the square of the density). As described above, when the frequency is lower than the commercial power frequency, the motor efficiency is lower than when the frequency is higher than the commercial power frequency, and the ratio of the motor loss to the theoretical compression work increases.

In addition, when operating at a low frequency, such as 28 Hz, which is lower than the commercial power frequency, the thrust bearing of the revolving type scroll compressor has a peculiar motion different from the thrust bearing which supports the general rotation motion called the revolving motion. Since the radius of rotation is the radius of revolution, the radius of rotation is smaller than that of a bearing that supports the radial load of the rotating shaft such as the main bearing, so the sliding speed is reduced, the lubrication performance is reduced, and the refrigerant compressor There is also an adverse effect on the reliability, such as the durability. The decrease in lubrication performance cannot maintain lubrication due to fluid lubrication,
A boundary lubrication state increases wear and the like, or, in the worst case, seizes and locks, which may cause a fatal problem for the refrigerant compressor that it becomes inoperable. Further, in the boundary lubrication state, the friction coefficient increases, which causes an increase in bearing loss, which is also a problem in performance.

As described above, in an environment near a room temperature set value where the operation time is relatively long when the refrigerant compressor is operated for a relatively long time, in order to obtain more comfort, the refrigerant compressor is operated at a low frequency lower than the commercial power supply at a low speed. In a refrigeration cycle device that performs variable-speed driving of a conventional refrigerant compressor that has been finely controlled, the performance (coefficient of performance) of the refrigerant compressor during such low-speed operation must be lower than the performance at frequencies higher than the commercial power supply. Therefore, the period energy consumption efficiency (hereinafter referred to as SEER; SEER = Seasonal Energy Efficiency Ratio) was low. Further, when the operation time at a low frequency lower than the commercial power supply is long, there is a problem that the durability of the refrigerant compressor is reduced due to the deterioration of the bearing performance.

SUMMARY OF THE INVENTION The present invention has been made to solve the above-described problems, and it is an object of the present invention to provide a refrigerant compressor that performs variable speed driving and has a high SEER and can further improve durability. SEE while gaining comfort
It is an object of the present invention to provide a refrigeration cycle apparatus that performs variable speed driving of a refrigerant compressor with high R and high reliability, and further provides a small and lightweight refrigerant compressor and a refrigeration cycle apparatus including the same. And

[0015]

According to a first aspect of the present invention, there is provided a refrigerant compressor used in a refrigeration cycle and performing variable speed driving by changing the frequency of a driving power supply, wherein the refrigerant compressor is Except at the time of start-up, it has a predetermined operating frequency range during a steady operation, and the lowest operating frequency in the operating frequency range is set to a frequency equal to or higher than a frequency near the commercial power supply frequency.

According to a second aspect of the present invention, in the first aspect, the refrigerant compressor is a scroll compressor in which a pair of spiral scroll teeth are combined to compress the working refrigerant. It was made.

According to a third aspect of the present invention, in the first or second aspect, a mixed refrigerant obtained by mixing difluoromethane (HFC32) and pentafluoroethane (HFC125) is used as a working refrigerant. It was what was.

According to a fourth aspect of the present invention, there is provided a refrigeration cycle apparatus in which a refrigerant compressor, a heat exchanger, an expansion mechanism, and the like are connected by piping. It has a predetermined operating frequency range, the lowest operating frequency in the operating frequency range is set to a frequency equal to or higher than the frequency near the commercial power supply frequency, and variable speed driving is performed by changing the frequency of the drive power supply.

According to a fifth aspect of the present invention, in the fourth aspect, the refrigerant compressor is a scroll compressor in which a pair of spiral scroll teeth are combined to compress the working refrigerant. It was made.

According to a sixth aspect of the present invention, in the fourth or fifth aspect, a mixed refrigerant obtained by mixing difluoromethane (HFC32) and pentafluoroethane (HFC125) is used as a working refrigerant. This was a refrigerating cycle device.

[0021]

BEST MODE FOR CARRYING OUT THE INVENTION

Embodiment 1 FIG. Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. FIG. 1 is a basic configuration diagram of a refrigeration cycle apparatus for performing variable speed driving of a refrigerant compressor according to an embodiment of the present invention. In the drawing, 100 is a refrigerant compressor, 101 is an outdoor heat exchanger, and 102 is indoor heat. Exchanger, 103 is a four-way valve,
104 is an outdoor expansion valve, 105 is an indoor expansion valve, and 106 is an accumulator, which are connected to form a refrigeration cycle. Reference numeral 107 denotes a commercial power supply.
7 is connected to a frequency variable device 108 such as an inverter circuit. The output of the frequency variable device 108 and the refrigerant compressor 100
Motors are connected. The current from the commercial power supply 107 is subjected to variable voltage variable frequency control by the frequency variable device 108,
The refrigerant compressor 100 is driven at a variable speed.

FIG. 2 is a sectional view showing a scroll compressor as the refrigerant compressor 100 in FIG. 1, and a revolving type scroll compressor in which a pair of spiral scroll teeth are combined to compress the working refrigerant. It is composed of The basic structure and operation of this refrigerant compressor are shown in FIGS.
Are the same as those of the scroll compressor shown in FIG.

In the above configuration, the refrigerant compressor 10
The lowest operating frequency in the steady operating frequency range excluding the start-up of 0 is a frequency near the frequency of the commercial power supply 107 (60 Hz or 50 Hz in Japan) or a frequency equal to or higher than the frequency of the commercial power supply 107. In addition, the steady operation frequency range is shifted to a higher frequency side as a whole than before. Specifically, here, the lowest operating frequency in a steady operating frequency range excluding the time of starting is 60 Hz, and the steady operating frequency range is 60 Hz to 240 Hz.
This means that the conventional steady operating frequency range is 30 Hz to 120 H
In this case, the stroke capacity of the refrigerant compressor 100 can be reduced by half to achieve the same capacity.
The shapes of the a and 22a can be reduced in size. When the spiral scroll teeth are formed from the involute curve, the stroke volume is calculated by the following equation. Vst = (2N-1) πp (p−2t) h where Vst: stroke volume N: number of turns p: pitch p = 2πA A: base circle radius of involute t: tooth thickness h: tooth height Number of turns N and If the tooth thickness t is not changed, the tooth height h and the pitch p can be reduced. 30Hz
As compared with the conventional refrigerant compressor in which the operating range is up to 120 Hz, the reduction of the tooth height h reduces the leak passage area formed by the radial clearance C between the side surfaces of the scroll teeth shown in FIG. In addition, the reduction of the pitch p can reduce the area of the leakage flow path formed by the axial clearance K between the tooth tip surface of the scroll tooth and the end surface of the base plate of the other scroll similarly shown in FIG. As a result, at the lowest frequency operation where the ratio of leakage loss to the theoretical compression work is high, that is, 30H
The lowest frequency of the refrigerant compressor of the embodiment in which the stroke volume of the conventional refrigerant compressor whose operating range is z to 120 Hz and the stroke volume of the conventional refrigerant compressor whose operating range is 60 Hz to 240 Hz is halved as described above. When operated at the same pressure condition at 60 Hz, the leakage loss is smaller in the embodiment because of the smaller leakage area despite the same capacity (same refrigeration capacity and the same theoretical compression work if the volume efficiency is the same). Becomes smaller. Further, in the embodiment, the amount of the working refrigerant in the compression chamber formed at the outermost periphery leaks to the outside of the compression chamber in the early stage of the compression stroke due to the decrease in the leak area, and the theoretical suction amount per unit time (frequency X stroke volume) is the same, so that the volume efficiency can be increased as compared with the conventional example. Since the volumetric efficiency is higher, the theoretical compression work increases, and the leakage loss decreases as described above. Therefore, the ratio of the leakage loss to the theoretical compression work in the embodiment can be reduced as compared with the conventional example, and the ratio of the leakage loss is particularly high. Example minimum frequency 30Hz
Compared to the compressor performance during operation, the lowest frequency of the embodiment 6
The compressor performance at 0 Hz operation is greatly improved. Of course, even at other frequencies, for the same reason, comparing the frequencies having the same capacity under the same pressure condition, the embodiment has a higher volumetric efficiency than the conventional example, and the ratio of the leakage loss to the theoretical compression work is higher. Can be lowered. The fact that the shape of both scroll teeth 21a and 22a can be reduced means that the fixed scroll 21 and the swivel scroll 22 can be downsized.
The main frame 26 for storing the scroll 22 and fixing the fixed scroll 21 can be downsized. The size reduction of the scrolls 21 and 22 is achieved by the thrust bearing 26.
a and the load acting on the main bearing 26b and the like can also be reduced.

The driving torque of the electric motor 28 required to compress the working refrigerant under the same pressure condition is a value obtained by dividing the theoretical compression work by (2π × frequency). Therefore, in this embodiment, the driving torque of the electric motor 28 required to compress the working refrigerant at the same theoretical compression work, that is, at the same refrigeration capacity under the same pressure condition, is twice as high in frequency, so that it may be half that of the conventional example. That is. Therefore, the electric motor 2
8 also enables a significant reduction in size compared to the prior art.
In the embodiment, since the operation at a low frequency lower than the commercial power supply frequency, which is a point at which the stall torque sharply decreases, is not performed within the steady operation range, the problem to be solved by the invention is described in the section of the problem to be solved by the invention. The aforementioned problem of motor loss does not occur. That is, even at the lowest frequency operation in the steady operation frequency range, the operation can be performed with the same magnetic flux density as at other frequencies, so that the iron loss can be reduced and the motor efficiency can be improved as compared with the conventional lowest frequency operation. Therefore, the ratio of the motor loss to the theoretical compression work can be reduced during the lowest frequency (60 Hz) operation of the embodiment compared to the conventional lowest operation frequency (30 Hz), and the compressor performance is improved. Here, if the same motor as the conventional one is used, ignoring the downsizing of the motor 28, the ratio of the operating torque to the stop torque becomes extremely low, and the motor efficiency can be greatly increased. Therefore, it is more advantageous to reduce the size. By setting the lowest frequency in the steady operation frequency range to a frequency near the commercial power supply frequency or a frequency equal to or higher than the frequency of the commercial power supply, the stop torque of the motor suddenly increases in the steady operation range. There is no point of lowering, so at the lowest frequency operation, motor efficiency such as copper loss increases due to an increase in current value, or iron loss increases by setting a high voltage. Factor can be eliminated. Therefore, the motor efficiency during the lowest frequency operation can be greatly increased, and the compressor performance during the lowest frequency operation can be greatly improved.

Next, the thrust bearing 26a will be described. In the embodiment, since the scrolls 21 and 22 can be downsized by halving the stroke volume as described above, the compression load in the thrust direction generated by the compression of the working refrigerant independent of the frequency is reduced by the pitch p. It is reduced. For this reason, the size of the thrust bearing 26a can be reduced. Since the lowest frequency in the steady operation range is twice that of the conventional example, the sliding speed increases. The orbital radius of the orbiting scroll 22 is calculated by the following equation: orbital radius Rc = p / 2-t, where: p: pitch p = 2πA A: base circle radius of involute t: tooth thickness Spiral shape due to half reduction in stroke volume Scroll teeth 21a, 22a
When downsizing is performed, in the embodiment, not only the pitch p is reduced, but also the tooth height h is also reduced, so that the revolving radius is not halved to the revolving radius. It is much larger than the value of half. Therefore, the sliding speed with respect to the thrust bearing 26a at the lowest operation frequency can be increased as compared with the conventional case. Therefore, the thrust bearing 26a
The lubrication performance of the thrust bearing 26a at the lowest operation frequency where the operation time is relatively long when compared to the conventional one can be improved as compared with the conventional case, and the fluid lubrication state is reliably assured and the durability of the refrigerant compressor is large. Be improved. In the fluid lubrication state, an increase in the sliding speed causes an increase in the sliding loss, and the higher the frequency, the higher the sliding loss in the thrust bearing. In the frequency range of the embodiment, there is a concern about an increase in thrust bearing sliding loss. However, since the rotation radius of a thrust bearing is originally small, the ratio of the sliding loss to the theoretical compression work is small if it is fluid-lubricated. Therefore, the sliding speed is increased at a high frequency such as 240 Hz in the embodiment, and the sliding loss of the thrust bearing 26a is increased as compared with the corresponding conventional high frequency such as 120Hz, but the compressor performance is not improved. The decline is small. On the other hand, at the time of the lowest frequency, by ensuring the fluid lubrication state, the friction coefficient can be reduced as compared with the conventional boundary lubrication state, and not only the reliability can be improved but also the sliding loss can be reduced.

For bearings that support the radial load of a relatively rotating shaft such as the main bearing 26b, the sub-bearing 31a, and the oscillating bearing 22d (collectively referred to as radial bearings), the radius of rotation is equal to the bearing radius. Since the radius itself is large, and the sliding speed is larger than that of the thrust bearing 26a by the rotation radius, sufficient lubrication performance is ensured even at the lowest operating frequency in the conventional example. Therefore, if the minimum oil film thickness of each radial bearing portion at the lowest operating frequency of the conventional example and the corresponding minimum oil film thickness of each bearing portion under the same pressure conditions at the lowest frequency of the embodiment are formed, Also in the embodiment, sufficient lubrication performance can be secured. In the same embodiment as described above, the scrolls 21 and 22 can be miniaturized by halving the stroke volume as compared with the conventional example. Therefore, the radial compression load generated by the compression of the working refrigerant independent of the frequency is particularly high in the tooth height. The effect of reducing h is reduced. The reduction in the pitch p also contributes to this. Further, since the orbiting scroll 22 for generating the centrifugal force is reduced in size or weight, and the orbital radius of the orbiting scroll 22 is reduced, the balancer 30a for balancing the entire rotating system against the centrifugal force generated by the orbital motion. ,
30b can be reduced in size and weight. In addition, the distance between the main bearing 26b and the sub-bearing 31a and the distance between the balancers 30a and 30b are reduced due to the miniaturization of the electric motor 28. Therefore, if the minimum oil film thickness of each radial bearing at the minimum operating frequency of 60 Hz in the embodiment is to be made equal to the minimum operating frequency of 30 Hz of the conventional example, the centrifugal force increases with the rotation speed, that is, the square of the frequency. , From the above reasons and because the frequency is low and the effect of centrifugal force is small,
The working radial load is reduced in the embodiment, and the frequency is doubled to increase the sliding speed, so that each radial bearing can be made smaller than the conventional example. If the bearing is reduced in size such that the bearing diameter is reduced by half, the sliding speed will not change even if the frequency is doubled, but since the load is reduced, the lubrication performance will be reduced even if such a reduction is made. There is no lowering than in the conventional example, and sufficient durability can be obtained. On the high frequency side where the influence of centrifugal force is large, the centrifugal force increases with the square of the frequency, so even if there is a radial load reducing element as described above, it is reversed from the low frequency side and acts on each radial bearing. The radial load will be higher than the corresponding conventional high frequency side. However, as each radial bearing is miniaturized, the coefficient of friction is reduced. In the embodiment, since each radial bearing is downsized mainly by reducing the diameter of the bearing, the radius of rotation is reduced, and even if the frequency is doubled, the sliding speed is almost increased. Not. By taking such measures, the sliding loss of the radial bearing on the high frequency side can be suppressed to be equal to or less than the sliding loss at the corresponding frequency in the conventional example. Therefore, the sliding loss of the radial bearing, which has a higher ratio to the theoretical compression work on the high frequency side than the conventional one, can be equal to or less than the ratio of the theoretical compression work on the high frequency side even if the frequency is doubled in the embodiment. it can.

FIG. 4 shows the revolving type scroll compressor used in the present embodiment and the revolving type scroll compressor having a steady operating frequency range of 30 to 120 Hz, which is mentioned as a conventional example, under the same conditions. 3 shows a comparison of the performance of the compressor alone. The vertical axis in the figure indicates the coefficient of performance COP (CoP) of the conventional compressor at 60 Hz.
efficient of performance) as 1 and the ratio to it. Further, the horizontal axis represents the frequency, and the corresponding frequencies (that is, the frequencies having the same capacity) in the conventional embodiment and the present embodiment are at the same position. As shown in FIG. 4, the COP at the lowest frequency is greatly improved for the above reason. The higher the frequency, the higher the C of the embodiment compared to the conventional example.
The superiority of the OP decreases, but this is because, when the corresponding frequencies are compared with each other, the advantage of reducing the leakage loss of the embodiment is that overshooting other than the thrust bearing 26a described above. This is because the increase is offset by an increase in secondary copper loss due to a discharge pressure loss called a loss or a skin effect of the electric motor 28. However, even at the highest frequency, the COP is higher in the embodiment. As for the overshoot loss, the higher the frequency, the higher the ratio to the theoretical compression work.
Although the embodiment has halved the stroke volume and the discharge volume flow rate has been halved, the effect of doubling the frequency is greater, and when comparing the corresponding frequencies, the embodiment is more effective than the theoretical compression work. The ratio of the overshoot loss becomes higher, and the higher the frequency, the higher the ratio as compared with the embodiment. Regarding the skin effect, this is a phenomenon in which the secondary resistance and the secondary inductance change due to the uneven distribution of the current flowing through the rotor 28b. The higher the frequency becomes, the more the secondary copper loss particularly increases. is there.

As described above, although the improvement in the compressor performance on the high frequency side is small, the minimum operating frequency in the steady operating frequency range other than the start-up period is 60 Hz, and the steady operating frequency range is 60 Hz to 240 Hz. By shifting to a higher frequency side, the compressor performance of the refrigerant compressor 100 at the time of the lowest frequency operation having a relatively long operation time can be greatly improved over the period, so that the SEER of the refrigerant compressor is improved. In addition, the SEER of the refrigeration cycle device that carries the variable speed drive and has it improved. By performing the operation at the lowest operation frequency in the operation frequency range, it is possible to perform fine control in an environment near room temperature setting as in the related art. Further, since the lubrication performance of the thrust bearing 26a during the lowest frequency operation can be improved, the durability of the refrigerant compressor 100 is improved, and the reliability is also improved. Furthermore, as described above, since each component can be reduced in size, the overall size and weight of the refrigerant compressor 100 can be reduced, and the cost can be reduced. In the scroll compressor used in FIG. 4, the embodiment has a compressor outer diameter of 88%,
Compressor height is reduced to 78% and compressor weight is 65%
And lighter. This is also the refrigerant compressor 100
This also means that the refrigeration cycle device on which the device is mounted can be reduced in size and weight and cost can be reduced.

The advantages of using a scroll compressor as the refrigerant compressor 100 will now be described. In a scroll compressor, the processes of suction, compression and discharge proceed simultaneously and continuously, and furthermore, a compression chamber of an intermediate pressure is formed between suction and discharge, so that the pressure rise speed is slow, so that torque fluctuation is very large. small. Since the discharge is a substantially continuous flow, the pressure fluctuation of the working refrigerant to be discharged is very small. Therefore, since the vibration is low, it is suitable for operation at a high frequency.
The operation at a high frequency is possible even if the steady operation frequency range is shifted to the higher frequency side as a whole as a frequency near the frequency of the above or a frequency equal to or higher than the frequency of the commercial power supply 107. It is. In addition, since a discharge valve is not necessarily required, problems such as noise, breakage, and valve followability due to the discharge valve during high-frequency operation do not occur.

In the above-described embodiment, the scroll compressor has a revolving type in which one scroll is stationary and the other scroll does not rotate around the center of the stationary scroll. Scroll compressors are used, but the scroll compressors are eccentrically combined, and both scrolls rotate in the same direction to compress the working refrigerant. There is a machine. FIG.
FIG. 6 is a longitudinal sectional view of the two-rotation type scroll compressor, and FIG. 6 is an explanatory view showing the compression principle of the two-rotation type scroll compressor, which is the same as the above-mentioned revolving type scroll compressor. Or, corresponding parts are denoted by the same reference numerals and description thereof will be omitted. In the drawing, reference numeral 41 denotes a drive scroll in which spiral scroll teeth 41a are erected on a base plate 41b, and a drive shaft 42 is connected to the opposite surface. Reference numeral 43 denotes a spiral scroll on a base plate 43b. The teeth 43a are erected, and the driven shaft 44
Are connected to each other to form a driven scroll.
1a and 43a are eccentrically combined. Electric motor 2
The drive shaft 42 is driven to rotate by 8, and the connected drive scroll 41 rotates by itself. When the driving scroll 41 rotates, the base plate 41 of the driving scroll 41 is rotated.
Since the drive pin 45 erected in b is fitted in the radially extending groove 46 formed in the driven scroll 43, the driven scroll 43 also rotates in the same direction as the drive scroll 41. I do. As the scrolls 41, 43 combined eccentrically rotate in the same direction as shown in FIG.
As shown in FIG. 3, a crescent-shaped compression chamber 25 formed between the scroll teeth 41a and 43a sequentially reduces the volume thereof, compresses the working refrigerant, and discharges it through the discharge hole 24. If the drive scroll 42 is formed so that the center of gravity of the drive scroll 41 is concentric with the drive shaft 42, no eccentric movement is performed on the drive bearing 47 that supports the drive shaft 42 in the radial direction, and the driven scroll is driven. If the driven scroll 43 is formed such that the center of gravity of the scroll 43 is concentric with the driven shaft 44, no eccentric movement is performed with respect to the driven bearing 48 that supports the driven shaft 44 in the radial direction, and thus the rotation is performed. Complete balancing of the entire system is built up and does not require the installation of balance weights.

Such a double-rotation type scroll compressor is used as the refrigerant compressor 100, and the lowest operating frequency in a steady operating frequency range except at the time of startup is a frequency near the frequency of the commercial power supply 107, Alternatively, if the frequency is equal to or higher than the frequency of the commercial power supply 107 and the steady operation frequency range is shifted to a higher frequency side as compared with the conventional one, the same effect as that of the above embodiment can be obtained, and Therefore, the vibration can be further reduced during high frequency operation.

When the refrigerant compressor 100 is stopped for a long period of time, a so-called stagnation phenomenon occurs in which the working refrigerant condenses in the closed container 27 and accumulates in a liquid state. In the case where the amount of the liquid refrigerant that has fallen in the scroll compressor is large, a state in which the liquid is filled even in the compression chamber 25 occurs, and the scroll compressor may be started from such a state. At this time, if it starts up at a high frequency, the volume change rate of the compression chamber with respect to time is large,
In particular, a sudden increase in the internal pressure, that is, a pressure pulse is generated in the intermediate pressure chamber, which may cause a problem that the spiral scroll teeth are damaged. Therefore, in this embodiment, the lowest frequency in the steady operation frequency range is a frequency near the frequency of the commercial power supply 107 or a frequency equal to or higher than the frequency of the commercial power supply 107.
A slow start from a low frequency such as 0 Hz is performed, and by increasing the frequency in a stepwise manner, generation of a pressure pulse at the time of start-up of sleep is prevented.

Embodiment 2 FIG. Considering the use of a rotary compressor as the refrigerant compressor 100, when the lowest frequency operation is performed, similar to the scroll compressor, the stroke volume is reduced, and especially the side clearance between the vane and the vane groove and the rolling are reduced. The reduction of the leakage area formed by the radial clearance between the outer peripheral surface of the piston and the inner peripheral surface of the cylinder can reduce the leakage loss, and the same effect can be obtained with respect to the motor loss, so that the compressor performance can be greatly improved. . However, unlike the scroll compressor, the rotary compressor performs the suction, compression and discharge strokes in one revolution, so that the pressure rise speed is large, and therefore the torque fluctuation is larger than that of the scroll compressor. Compressors are not as suitable for high-speed operation as scroll compressors because they require a discharge valve. Therefore, the high frequency side is restricted as compared with the scroll compressor, and the operating frequency range is narrowed as compared with the scroll compressor. It is considered that the greatest factor hindering the high-speed operation of the rotary compressor is the sliding of the vane and the outer peripheral surface of the rolling piston. When the frequency operation is performed, wear of the vane tip and the outer peripheral surface of the rolling piston becomes severe. However, recently, there has been proposed a rotary compressor in which a vane and a rolling piston are integrally formed and an outer peripheral surface of the vane and the rolling piston are non-sliding. The problem of wear has been eliminated.

Embodiment 3 In recent years, from the viewpoint of environmental protection,
Hydrofluorocarbon-based refrigerants (hereinafter referred to as HFC-based refrigerants) that do not contain chlorine atoms in the molecular structure that destroy the ozone layer are being used as working refrigerants for refrigeration cycle devices and are currently mainly used for air conditioning. Chlorodifluoromethane (hereinafter referred to as HCFC22) has been determined to be completely abolished in the near future. Difluoromethane (hereinafter, referred to as HFC32) as an alternative refrigerant to this HCFC22,
Pentafluoroethane (hereinafter referred to as HFC125), 1,
1,1,2 tetrafluoroethane (hereinafter referred to as HFC134a), 1,1,1 trifluoroethane (hereinafter HFC143a)
), 1,1 difluoroethane (hereinafter HFC152a)
HFC-based refrigerants such as a single refrigerant or a plurality of mixed refrigerants. HFC32, HFC125 and HFC134a for air conditioning
HFC32 / 125 / 134a refrigerant mixed with seeds, or HFC3 mixed with HFC32 and HFC125
The 2/125 refrigerant is considered the most promising. Especially HF
C32 / 125 is a high-pressure refrigerant and has a considerably higher saturation pressure at the same temperature than HCFC22. Therefore, under the same temperature conditions as before, the difference between the high and low pressures is higher than that of HCFC22.
Extremely large. On the other hand, since the density is higher than that of the HCFC 22, the stroke volume may be smaller than that of the HCFC 22 in order to obtain the same refrigerating capacity as that of the HCFC 22.

When this HFC32 / 125 is used as a working refrigerant of a refrigeration cycle apparatus for performing variable speed driving of a refrigerant compressor, as described above, the pressure difference between the HCFC22 and the high and low pressures increases, so that the leakage loss increases, and especially the leakage loss increases. Compressor performance at the lowest frequency operation, such as the conventional 30 Hz where loss is high relative to theoretical compression work, is significantly reduced. Therefore, the application of the present invention, that is, the refrigerant compressor 100
The minimum operating frequency in a steady operating frequency range other than the start-up period is set to a frequency near the frequency of the commercial power supply 107 or a frequency equal to or higher than the frequency of the commercial power supply 107. If the frequency range is shifted to the higher frequency side as compared with the conventional one, the effect of reducing the leakage loss is particularly effective for the above-mentioned reason, and the compressor performance at the lowest frequency operation is greatly improved. Further, since the stroke volume can be reduced from the time of HCFC 22:00, the discharge volume flow rate at the time of high frequency operation is reduced, and the overshoot loss can be reduced. Therefore, the compressor performance can be improved even on the high frequency side,
Compressor performance over the entire operating frequency range In other words, C
The OP can be made almost flat. Therefore, the SEER of the refrigerant compressor 100 is improved, and the SEER of the refrigeration cycle device that carries the refrigerant compressor and performs variable speed driving can also be improved.

In the embodiment of the present invention, the predetermined operating frequency range during steady operation is 60 Hz to 240 Hz.
However, the predetermined operating frequency range during steady operation of the present invention is 60 Hz to 240 Hz.
Hz, the minimum operating frequency must be a frequency higher than the frequency near the commercial power supply frequency, that is, a frequency higher than the point where the stall torque suddenly decreases as described in the subject column. Also, at the maximum operating frequency, there is a problem that performance is deteriorated due to an increase in loss such as sliding loss and overshoot loss of the bearing portion at high speed, or secondary copper loss due to the skin effect of the motor. , And can be appropriately selected between these two frequencies. According to the present invention, by setting the predetermined operating frequency range during the steady operation of the refrigerant compressor as described above, the size and weight are reduced, and the SEER is high.
Further, it is possible to provide a low-cost refrigerant compressor that performs variable speed driving and can also improve durability. In addition, a refrigeration cycle apparatus that can be reduced in size and weight and has a high SEER while obtaining sufficient comfort, and that performs variable speed driving of a refrigerant compressor with high reliability can be provided at low cost.

[0037]

As described above, according to the first aspect of the present invention, in a refrigerant compressor that performs variable speed driving by changing the frequency of a driving power supply, the refrigerant compressor operates in a steady state except at startup. Sometimes has a predetermined operating frequency range, the lowest operating frequency of the operating frequency range is set to a frequency equal to or higher than the frequency near the commercial power supply frequency, the theory of the refrigerant compressor during the lowest frequency operation in the steady operating frequency range The ratio of the leakage loss and the motor loss to the compression work can be reduced, and the compressor performance of the refrigerant compressor during the lowest frequency operation, which has a relatively long operation time over a period, can be greatly improved, so that the refrigerant compression SEE
R can be improved. By performing the operation of the refrigerant compressor at the lowest operation frequency in the operation frequency range, fine control can be performed in an environment near room temperature setting as in the related art. Further, the size and weight of the compressor can be reduced, and the cost can be reduced.

According to a second aspect of the present invention, in the first aspect, the refrigerant compressor is a scroll compressor that compresses the working refrigerant by combining a pair of spiral scroll teeth. Therefore, in addition to the effects of the first invention, the lubrication performance of the bearing portion of the compressor during the lowest frequency operation can be improved, and the durability and reliability of the refrigerant compressor can be improved. Further, even in high frequency operation, a pressure fluctuation is small, and a compressor with low vibration can be obtained.

In the third aspect of the present invention, a mixed refrigerant of HFC32 and HFC125 is used as the working refrigerant for the compressor whose minimum operating frequency is equal to or higher than the frequency near the commercial power supply frequency. The increase in leakage loss is reduced. In addition, HFC is used as the working refrigerant.
By using a refrigerant mixture of HFC32 and HFC125, the compressor performance on the high frequency side can be improved, the performance difference of the refrigerant compressor depending on the frequency can be corrected, and the performance of the compressor becomes almost flat over the entire operating frequency range. The SEER of the refrigerant compressor can be further improved.

According to a fourth aspect of the present invention, in the refrigeration cycle apparatus, the refrigerant compressor has a predetermined operating frequency range during a steady operation except for a start-up operation, and sets a minimum operating frequency of the operating frequency range to a minimum operating frequency. Since the frequency is set to a frequency equal to or higher than the frequency near the commercial power supply and the driving power supply is changed to perform variable speed driving, the SEER of the refrigeration cycle apparatus can be improved. Further, by mounting the compact and lightweight refrigerant compressor, the size and weight of the refrigeration cycle device and the cost can be reduced.

Further, in the fifth invention, since the scroll compressor is used as the refrigerant compressor in the fourth invention, the durability and the reliability are improved, and the pressure fluctuation even in high frequency operation. By using a small, low-vibration compressor,
A highly reliable refrigeration cycle device with improved performance can be obtained.

According to a sixth aspect of the present invention, in the fourth or fifth aspect, difluoromethane (HFC32) and pentafluoroethane (HFC12) are used as working refrigerants.
Refrigeration cycle apparatus using a mixed refrigerant obtained by mixing 5) with the above. Therefore, the increase in leakage loss due to the use of high-pressure refrigerant is mitigated, and the performance is improved by using a refrigerant compressor with improved SEER. Thus, a refrigeration cycle apparatus with improved SEER can be obtained.

[Brief description of the drawings]

FIG. 1 is a basic configuration diagram of a refrigeration cycle apparatus that performs variable speed driving of a refrigerant compressor according to an embodiment of the present invention.

FIG. 2 is a longitudinal sectional view showing a configuration of a refrigerant compressor (revolution type scroll compressor) according to an embodiment of the present invention.

FIG. 3 shows a scroll according to an embodiment of the present invention.
FIG. 4 is an explanatory view of a compression chamber leakage clearance of a compressor.

FIG. 4 is a single unit performance comparison diagram of an embodiment of the present invention and a conventional refrigerant compressor (revolution type scroll compressor).

FIG. 5 is a longitudinal sectional view showing a configuration of a refrigerant compressor (double-rotating scroll compressor) according to an embodiment of the present invention.

FIG. 6 is an explanatory diagram showing the compression principle of a double-rotating scroll compressor.

FIG. 7 is a longitudinal sectional view showing a configuration of a conventional refrigerant compressor (rotary compressor).

FIG. 8 is a cross-sectional view of a compression element of the rotary compressor.

FIG. 9 is an explanatory diagram showing the compression principle of a revolution type scroll compressor.

FIG. 10 is a longitudinal sectional view showing a configuration of a conventional refrigerant compressor (revolution type scroll compressor).

[Explanation of symbols]

21a spiral scroll teeth (fixed scroll teeth),
22a spiral scroll teeth (orbiting scroll teeth),
REFERENCE SIGNS LIST 100 refrigerant compressor (scroll compressor), 101 heat exchanger, 102 heat exchanger, 103 switching valve, 104
Expansion valve, 105 Expansion valve, 107 Drive power supply.

 ──────────────────────────────────────────────────続 き Continuing from the front page (72) Eiji Watanabe, 2-3-2 Marunouchi, Chiyoda-ku, Tokyo In-house Mitsui Electric Co., Ltd. (72) Inventor Shin Sekiya 2-3-2, Marunouchi, Chiyoda-ku, Tokyo Rishi Electric Co., Ltd. (72) Inventor Toshiyuki Nakamura 2-3-2 Marunouchi, Chiyoda-ku, Tokyo Mitsui Electric Co., Ltd. (72) Inventor Kenji Suzuki 2-3-2 Marunouchi, Chiyoda-ku, Tokyo Mitsui Electric Inside the corporation

Claims (6)

[Claims] 1. A refrigerant compressor used in a refrigeration cycle and performing variable speed driving by changing the frequency of a driving power supply, wherein the refrigerant compressor operates at a predetermined operation during a steady operation except for a start-up operation. A refrigerant compressor having a frequency range, wherein the lowest operating frequency in the operating frequency range is set to a frequency equal to or higher than a frequency near a commercial power frequency. 2. The refrigerant compressor according to claim 1, wherein the refrigerant compressor is a scroll compressor in which a pair of spiral scroll teeth are combined to compress the working refrigerant. 3. The working refrigerant is difluoromethane (HF).
The refrigerant compressor according to claim 1 or 2, wherein a mixed refrigerant obtained by mixing C32) and pentafluoroethane (HFC125) is used. 4. In a refrigeration cycle apparatus in which a refrigerant compressor, a heat exchanger, an expansion mechanism, and the like are connected by piping, the refrigerant compressor has a predetermined operating frequency range during a steady operation except for a start-up operation. A refrigeration cycle apparatus characterized in that the refrigeration cycle apparatus is driven at a variable speed by setting the lowest operation frequency in the operation frequency range to a frequency equal to or higher than the frequency near the commercial power supply frequency and changing the frequency of the drive power supply. 5. The refrigeration cycle apparatus according to claim 4, wherein the refrigerant compressor is a scroll compressor in which a pair of spiral scroll teeth are combined to compress the working refrigerant. 6. Difluoromethane (HF) as a working refrigerant
The refrigeration cycle apparatus according to claim 4, wherein a mixed refrigerant obtained by mixing C32) with pentafluoroethane (HFC125) is used.