GB2315299A

GB2315299A - Variable speed drive refrigerant compressor

Info

Publication number: GB2315299A
Application number: GB9714547A
Authority: GB
Inventors: Eiji Watanabe; Shin Sekiya; Toshiyuki Nakamura; Minoru Ishii; Hiroshi Ogawa; Kazuyuki Akiyama; Kenji Suzuki
Original assignee: Mitsubishi Electric Corp
Current assignee: Mitsubishi Electric Corp
Priority date: 1996-07-11
Filing date: 1997-07-10
Publication date: 1998-01-28
Anticipated expiration: 2017-07-10
Also published as: CN1174294A; JPH1026425A; ITMI971600A1; GB9714547D0; IT1292487B1; GB2315299B; KR980010247A; ITMI971600A0

Abstract

In a refrigerant compressor which is used in a refrigeration cycle and carries out variable speed drive by changing a frequency supplied from a power source, the compressor has a predetermined operational frequency range at steady operation except for at a start-up, and the operational frequency range has a minimum operational frequency which is not less than a frequency near to the frequency of the power supply.

Description

VARIABLE SPEED DRIVE REFRIGERANT COMPRESSOR, AND REFRIGERATION CYCLE APPARATUS INCLUDING THE SAME The present invention relates to a refrigerant compressor which is used in a refrigeration cycle and carries out variable speed drive by changing a frequency supplied from a powers source by use of e.g. an inverter, and a refrigeration cycle apparatus which controls the operational frequency for a refrigerant compressor by use of e.g. an inverter to cause the compressor to carry out variable speed drive.

In a refrigeration cycle apparatus which causes a refrigerant compressor to carry out variable speed drive by use of e.g. an inverter, the compressor has a minimum operational frequency of about 28 Hz as described in JP A-7294030, which is a frequency lower than the frequency of a commercial power supply (60 Hz or 50 Hz in Japan).

As the refrigerant compressor in the refrigeration cycle apparatus which causes the compressor to carry out variable speed drive, a rotary compressor as shown in Figure 7 has been generally used, or a scroll compressor as shown in Figure 10 has been commonly used in recent years.

In Figure 7, there is shown a vertical crosssectional view of an example of the rotary compressor, depicting the schematic structure thereof. In Figure 8, there is shown a horizontal cross-sectional view of the compressing element shown in Figure 7. A hermetic casing 1 houses the compressing elements 2, and a variable speed compatible electric motor 3 which subjects the compressing elements 2 to variable speed drive through a crankshaft 4. Reference numeral 5 designates an intake tube which takes a working refrigerant into the compressing elements 2, and which communicates with a cylinder 6 with a compression chamber 11 formed therein.

The cylinder 6 have an intake hole 7 formed therein to connect between the intake tube 5 and a low pressure compartment lla in the compression chamber 11 as an inner space of the cylinder 6 as shown in Figure 8. The cylinder 6 has a rolling piston 8 arranged therein so as to be driven by rotation of the crankshaft 4 and be slidable along an inner side wall of the cylinder 6.

Reference numeral 9 designates a vane which is constantly urged against an outer peripheral surface of the rolling piston 8, which is arranged in a vane groove 10 formed in the cylinder 6 so as to be capable of reciprocating * ia radial direction of the cylinder, and which divides the compression chamber into a high pressure compartment llb and the low pressure compartment lla. The cylinder has c portion near to the vane 9 remote from the intake tube 7 formed with a discharge hole 12, which communicates with the high pressure compartment llb, discharges the compressed working refrigerant in the hermetic casing 1, and is opened and closed by a discharge valve 13.

In the rotary compressor thus constructed, the working refrigerant which has been taken in the low pressure compartment lla from the intake tube 5 through the intake hole 7 is gradually compressed to a predetermined pressure by eccentric rotary movement of the rolling piston 8 driven by the crankshaft 4, and is discharged from the high pressure compartment llb into the hermetic casing 1 through the discharge hole 12, opening the discharge valve 13.

In Figure 9, there is shown the basic components and the compression principle of an orbiting scroll compressor wherein one of scroll members is fixed, the other scroll member orbits about the center of the fixed scroll member without rotation to compress the working refrigerant. In Figure 10, reference numeral 21 designates a fixed scroll which is fixed with respect to a space, and which has one of the scroll members 21a.

Reference numeral 22 designates an orbiting scroll which carries out orbiting motion, and which has the other scroll member 22a formed in a spiral direction opposite to that of the fixed scroll member 21a and combined with the fixed scroll member 21a with a phase shift of 1800.

These scroll members 21a and 22a have a shape which is obtained by combining e.g. an involute curve or an arc.

Reference numeral 23 designates intake chambers, reference numeral 24 designates a discharge hole, and reference numeral 25 designates a compression chamber.

Now, the operation of the arrangement shown in Figure 9 will be explained. The orbiting scroll 22 carries out orbiting movement, that is to say revolving movement about the center of the fixed scroll member 21a without rotation, i.e. without changing its posture, thereby to take positions at o, 900, 1800 and 2700 shown in Figures 9(a), 9(b), 9(c) and 9(d) one after another. At 0. shown in Figure 9(a), trapping the working refrigerant in the intake chamber 23 is completed, and the compression chamber 25 is defined in a crescent form between the fixed scroll member 21a and the orbiting scroll member 22a. As the orbiting scroll 22 carries out the orbiting movement, the volume of the compression chamber 25 is gradually decreased to compress the working refrigerant until a location near to the center of the fixed scroll 21, and the compression chamber discharges the working refrigerant from the discharge hole 24.

Next, the specific structure and the operation of the scroll compressor will be explained. In Figure 10, there is shown a vertical cross-sectional view of an example of the orbiting scroll compressor. Reference numeral 21 designates the fixed scroll which has a base plate 21b formed with the scroll member 21a, and which is fixed to a main frame 26. Reference numeral 22 designates the orbiting scroll which has a base plate 22b formed with the scroll member 22a combined with the fixed scroll member 21a. An end surface of the base plate 22b remote from the scroll member 22a is formed with an orbiting bearing 22d. A compression load which acts on the orbiting scroll 22 and a thrust surface 22c receiving a thrust force due to the compression load applied to the orbiting scroll 22, and a radial load caused by a centrifugal force derived from the orbiting movement are born by the orbiting bearing 22d. The main frame 26 is fixed to a hermetic casing 27, has a thrust bearing 26a formed thereon, and supports the thrust surface 22c of the orbiting scroll 22 in an axial direction so that the orbiting scroll is slidable on the main frame. Reference numeral 28 designates a variable speed compatible electric motor which rotates a main shaft 29 in a variable speed design, and which is constituted by a stator 28a and a rotor 28b. The rotor 28b is fixed to the main shaft 29 by shrinkage fit, and has an upper portion and a lower portion formed with balancers 30a and 30b, which balance the overall rotation system against the centrifugal force derived from the orbiting movement of the orbiting scroll 22. The main shaft 29 is driven by the electric motor 28, and has opposite ends at both ends of the electric motor 28 supported in a radial direction by a main bearing 26b provided on the main frame 26 and by a sub bearing 31a provided on a sub frame 31. Reference numeral 32 designates an Oldham's ring which carries out repeatable linear motion with respect to both of the orbiting scroll 22 and the main frame 26.

The main shaft 29 has an upper end formed with an eccentric shaft 29a. By rotation of the main shaft 29, the eccentric shaft 29a transmits rotary movement to the orbiting scroll 22 through the orbiting bearing 22d.

Because the repeatable linear motion of the Oldham's ring restrains the rotation of the orbiting scroll 22, the orbiting scroll carries out orbiting movement with respect to the fixed scroll 21 to compress the working refrigerant as stated earlier. The hermetic casing 27 includes a lubricating oil reservoir 33 at a bottom thereof. The main shaft 29 has a lower end provided with a positive displacement pump 34 to supply a lubricating oil to the bearings 22d, 26a, 26b and 31a, and other sliding parts through an oil passage 29b formed in the main shaft 29.

When a refrigerant compressor such as the rotary compressor and the scroll compressor as stated earlier is operated at a low frequency less than the frequency of a commercial power supply such as 28 Hz as described in JP A-7294030 in a steady operational range except for at a start-up, there is created a problem in that the refrigerant compressor has performance extremely lowered in comparison with when the compressor is operated at the frequency of the commercial power supply or at a frequency higher than the frequency of the commercial power supply.

A first reason is that ratio of leakage loss to theoretical compression work increases at low speed operation. This is because the internal leakage of the working refrigerant per unit time which leaks into a compression chamber on a lower pressure side through several clearances during compression or after completion of compression is irrelevant to the speed (i.e. the operational frequency). Since a lower operation requires more time required for one rotation, the leakage becomes larger, and the ratio of leakage loss to theoretical compression work therefore increases under the same pressure condition to deteriorate coefficient of performance. With respect to the rotary compressor, examples of the clearances causing leakage are a lateral clearance between the vane and the vane groove and an axial clearance on an upper end surface of the rolling piston as a path through which the working refrigerant discharged into the hermetic casing at a discharge pressure leaks to both the high pressure compartment and the low pressure compartment, and also a radial clearance between an outer peripheral surface of the rolling piston and an inner peripheral surface of the cylinder and an axial clearance on an upper end surface of the vane as a path through which the working refrigerant leaks from the high pressure compartment into the low pressure compartment with respect to the rotary compressor. With respect to the rotary compressor, examples of the clearances are an axial clearance between leading end surfaces of the scroll members and the base plates of the opposed scrolls and a radial clearance between side surfaces of both scroll members as a path through which the working refrigerant leaks from a compression chamber into a compression chamber next thereto and having a lower pressure therein. In addition, in the scroll compressor, the working refrigerant in the compression chamber defined at an outermost peripheral portion could leak out of the compression chamber at the outermost peripheral portion through those paths at an initial stage during compression stroke, thereby extremely lowering volumetric efficiency at the low speed operation.

A second reason is that ratio of motor loss to theoretical compression work is high at low speed operation. The electric motor of the refrigerant compressor which carries out variable speed drive by changing the frequency supplied from a power source by use of e.g. an inverter is usually constituted by a three-phase squirrel-cage induction motor. In such a three-phase squirrel-cage induction motor, the frequency must be proportioned to the voltage in order to stabilize the magnetic flux density of flat rolled magnetic steel sheets and strip constituting the stator and the rotor, or the magnetic flux density in the clearance between the stator and the rotor regardless of frequencies. The lower the frequency is, the smaller stalling torque is obtained. In particular, there is a point where the stalling torque rapidly lowers at a lower frequency than the frequency of a commercial power supply. When the compressor is forced to be driven under the same pressure condition at a frequency higher than that of the commercial power supply, the operating torque approaches the stalling torque as the frequency (the speed) becomes lower. In that case, the current rises to increase copper loss (copper loss is proportional to the square of current), lowering motor efficiency. This means that the ratio of motor loss to theoretical compression work is raised. For these reasons, there is adopted a way to set the voltage higher than the voltage decided by the proportional relation in order to increase the stalling torque at a lower frequency than that of the commercial power supply, at which the stalling torque is small.

Although such a way can increase motor efficiency, the increased voltage rises the magnetic flux density of the flat rolled magnetic steel sheets and strip constituting the stator and the rotor, and the magnetic flux density in the clearance between the stator and the rotor to increase iron loss (iron loss is proportional to the square of magnetic flux density). When the compressor is driven at such a lower frequency than that of the commercial power supply, the motor efficiency is lowered in comparison with operation at a higher frequency than that of the commercial power supply, and the ratio of motor loss to theoretical compression work is raised as explained.

When the compressor is driven at a lower frequency, e.g. 28 Hz, than the commercial power supply frequency, in particular the thrust bearing of the orbiting scroll compressor is subjected to orbital motion which is peculiar motion different from the movement applied to a usual thrust bearing for bearing rotational motion.

Since the radius of revolution of the thrust bearing is the radius of the orbiting motion, the radius of revolution is smaller than the main bearing and other bearings for bearing a radial load of a rotational shaft.

This means that the sliding speed of the thrust bearing becomes small to lower lubricating performance, badly affecting the reliability of the compressor such as durability. The lowered lubricating performance fails to keep fluid lubrication state, and the fluid lubrication state shifts to boundary lubrication state, increasing e.g. wear, or causing seizing to lock the compressor and make it inoperative at the worst, which is fatal trouble for the compressor. In the boundary lubrication state, coefficient of friction also rises to introduce an increase in bearing loss, creating a problem in terms of performance.

The conventional refrigeration cycle apparatus causes the compressor to carry out variable speed drive with fine control by low speed operation at a lower frequency than the commercial power supply frequency in order to obtain a more comfortable atmosphere under such conditions that the compressor is driven at a set temperature near to a room temperature for a relatively long time in all seasons. As stated earlier, the compressor provides lower performance (coefficient of performance) in such low speed operation in comparison with the performance obtained in the operation at a frequency not less than the commercial power supply frequency. As a result, seasonal energy efficiency ratio (hereinbelow, referred to as "SEER") is low. Long operation at a lower frequency than the commercial power supply frequency creates a problem in that the durability of the compressor is lowered due to deteriorated bearing performance.

It is an object of the present invention to solve these problems, and to provide a refrigerant compressor capable of carrying out variable speed drive with increased SEER and improved durability, to provide a refrigeration cycle apparatus capable of providing a refrigerant compressor with variable speed drive with increased SEER and reliability as well as provision of sufficient comfort, and to provide a compact and lightweight refrigerant compressor and a refrigeration cycle apparatus with the compressor.

According to a first aspect of the present invention, there is provided a refrigerant compressor which is used in a refrigeration cycle and carries out variable speed drive by changing a frequency supplied from a power source, comprising the compressor having a predetermined operational frequency range at steady operation except for at a start-up, and the operational frequency range having a minimum operational frequency which is not lower than a frequency near to a commercial power supply frequency.

According to a second aspect of the present invention, the compressor is a scroll compressor which is constituted by combining a pair of scroll members so as to compress a working refrigerant in the first aspect.

According to a third aspect of the present invention, the working refrigerant is mixture of difluoromethane (HFC32) and pentafluoroethane (HFC125) in the first aspect or the second aspect.

According to a fourth aspect of the present invention, there is provided a refrigeration cycle apparatus which has a refrigerant compressor, heat exchangers and an expansion device connected by piping, comprising the compressor having a predetermined operational frequency range at steady operation except for at a start-up, the operational frequency range having a minimum operational frequency which is not lower than a frequency near to a commercial power supply frequency, and the compressor carrying out variable speed drive by changing a frequency supplied from a power source.

According to a fifth aspect of the present invention, the compressor is a scroll compressor which is constituted by combining a pair of scroll members so as to compress a working refrigerant in the fourth aspect.

According to a sixth aspect of the present invention, the working refrigerant is a mixture of difluoromethane (HFC32) and pentafluoroethane (EFCl45) in the fourth aspect or the fifth aspect.

In accordance with the first aspect, the ratio of leakage loss to theoretical compression work and the ratio of motor loss to theoretical compression work of the compressor at the minimum frequency operation in the steady operational frequency range can be reduced to remarkably improve the performance of the compressor at the minimum frequency operation which is made for a relatively long time in all seasons. As a result, the compressor can improve SEER. It is possible to carry out fine control at a set temperature near to a room temperature as usual by driving the compressor at the minimum operational frequency in the operational frequency range. It is also possible to make the compressor compact and lightweight and to reduce cost.

In accordance with the second aspect, it is possible to improve the lubricating performance at the bearing parts of the compressor at the minimum frequency operation in addition to the advantages offered by the first aspect. As a result, the compressor can improve durability and reliability. The compressor can minimize variations in pressure even at high frequency operation to offer a low vibration design.

In accordance with the third aspect, it is possible to alleviate an increase in leakage loss due to use of in particular a highly pressurized refrigerant. In addition, use of a mixture of BFC32 and iFCl25 as the working refrigerant can improve compressor performance on high frequency side, minimize differences in compression performance due to different frequencies, make compressor performance come closer to a flat manner, and further improve the SEER of the compressor.

In accordance with the fourth aspect, the refrigeration cycle apparatus can improve SEER. It is possible to make the refrigeration cycle apparatus compact and lightweight and to reduce cost by providing the compact and lightweight refrigerant compressor with the apparatus.

In accordance with the fifth aspect, it is possible to improve durability and reliability. Variations in pressure can be minimized even at high frequency operation. Use of the compressor with low vibration design can provide a refrigeration cycle apparatus with high reliability and improved performance.

In accordance with the sixth aspect, it is possible to alleviate an increase in leakage loss due to use of a highly pressurized refrigerant. Use of the refrigerant compressor with improved SEER can provide the refrigeration cycle apparatus with improved performance and SEER.

The invention will be further described by way of non-limitative example with reference to the accompanying drawings, in which: Figure 1 is a basic diagram of the refrigeration cycle apparatus wherein a refrigerant compressor carries out variable speed drive in accordance with a first embodiment of the present invention; Figure 2 is a vertical cross-sectional view of the structure of the refrigerant compressor (orbiting scroll compressor) according to the first embodiment; Figure 3 is a schematic view explaining leakage clearances at a compression chamber of the scroll compressor according to the first embodiment; Figure 4 is a graph showing comparison between the first embodiment and a conventional refrigerant compressor (orbiting scroll compressor) in terms of performance; Figure 5 is a vertical cross-sectional view of the structure of the refrigerant compressor (co-rotating) in accordance with the first embodiment; Figure 6 is a schematic view showing the compression principle of such a co-rotating; Figure 7 is a vertical cross-sectional view of the structure of a conventional refrigerant compressor (rotary compressor); Figure 8 is a horizontal cross-sectional view of the compression elements of the rotary compressor; Figure 9 is a schematic view showing the compression principle of an orbiting scroll compressor; and Figure 10 is a vertical cross-sectional view of the structure of a conventional refrigerant compressor (orbiting scroll compressor).

Now, Embodiments of the present invention will be described in detail referring to the accompanying drawings.

EMBODIMENT 1 In Figure 1, there is shown a basic diagram of a refrigeration cycle apparatus wherein a refrigerant compressor carries out variable speed drive in accordance with a first embodiment of the present invention. In this Figure, reference numeral 100 designates the refrigerant compressor, reference numeral 101 designates an outdoor heat exchanger, reference numeral 102 designates an indoor heat exchanger, reference numeral 103 designates a four port valve, reference numeral 104 designates an outdoor expansion valve, reference numeral 105 designates an indoor expansion valve, and reference numeral 106 designates an accumulator. These elements are connected to provide a refrigeration cycle.

Reference numeral 107 designates a commercial power supply, to which a variable frequency generating device 108 such as an inverter circuit is connected. The variable frequency generating device 108 has output connected to an electric motor in the refrigerant compressor 100. The current from the commercial power supply 107 is subjected to variable voltage variable frequency control by the variable frequency generating device 108, causing the compressor 100 to carry out variable speed drive.

In Figure 2, there is shown a cross-sectional view of a scroll compressor as the refrigerant compressor 100, which is an orbital scroll compressor which has a pair of scroll members combined to compress a working refrigerant. The basic structure and operation of the refrigerant compressor are the same as those of the scroll compressor shown in Figures 9 and 10. Identical or similar parts are indicated by the same reference numerals as those of the conventional compressor, and explanation of those parts will be omitted.

The refrigerant compressor 100 thus constructed has a minimum operational frequency in a steady operational frequency region except for at a start-up set at a frequency near to the frequency of the commercial power supply 107 (60 Hz or 50 Hz in Japan) or at a frequency not less than the frequency of the commercial power supply 107. The compressor has the entire operational frequency range shifted toward a high frequency side in comparison with the conventional compressor.

Specifically, the minimum operational frequency is set to 60 Hz in the steady operational frequency range except for the start-up, and the steady operational frequency range is set to 60 Hz - 240 Hz in the embodiment. This means that the compressor which has had the steady operational frequency range set to 30 Hz - 120 Hz is subjected to double speed operation, and that the compressor can provide the same volume with the swept volume of the refrigerant compressor 100 halved, allowing the size of both scroll members 21a and 22a to be made smaller. When the scroll members are in an involute curve form, the swept volume is found by the following equation: Vst = (2N-l)itp(p-2t)h wherein Vst represents the swept volume, N represents the number of turns, p represents pitch (p=2rA, A represents the radius of the involute base circle), t represents the thickness of the scroll members and h represents the height of the scroll members.

If the number of turns N and the thickness of the scroll members t are unchanged, the height of the scroll members h and the pitch p can be reduced. In comparison with the conventional refrigerant compressor having an operational range of 30 Hz - 120 Hz, a decrease in the height of the scroll members h can reduce the area of leakage paths defined by a radial clearance C between side surfaces of both scroll members as shown in Figure 3, and a decrease in the pitch p can also reduce the area of leakage paths defined by axial clearances K between leading edge surfaces of the scroll members and end surfaces of base plates of combined scrolls as shown in Figure 3. When the conventional refrigerant compressor having the operational range of 30 Hz - 120 Hz is driven at 30 Hz, i.e. the minimum frequency operation with the ratio of leakage loss to theoretical compressor work raised, and when the compressor according to the embodiment having the operational range of 60 Hz - 240 Hz with the swept volume halved in comparison with the conventional compressor as stated earlier is driven at 60 Hz as the minimum frequency under the same pressure condition as the conventional compressor, the compressor according to the embodiment can reduce leakage loss by a decrease in the leakage areas in spite of having the same volume (if the volumetric efficiency is unchanged, the same refrigerating capacity and the same theoretical compression work can obtained). In the embodiment, the decreased leakage areas can minimize the amount at which the working refrigerant in a compression chamber defined at an outermost peripheral portion leaks out of the compression chamber at an initial stage during compression stroke. The compressor according to the embodiment can raise volumetric efficiency in comparison with the conventional compressor because theoretical suction volume per unit time (frequency x swept volume) is the same as the conventional compressor. The theoretical compression work can increase by the raise in volumetric efficiency, and the leakage loss is lowered as explained. In accordance with the embodiment, the ratio of leakage loss to theoretical compression work can be reduced in comparison with the conventional compressor, and the compression performance of the embodiment under operation at 60 Hz as the minimum frequency can be remarkably improved in comparison with the conventional compressor performance under operation at 30 Hz as the minimum frequency wherein the ratio of leakage loss has been significantly high. With regard to operation at other frequencies, the compressor according to the embodiment can raise the volumetric efficiency and reduce the ratio of leakage loss to theoretical compression work in comparison with the conventional compressor for the same reasons as long as the compressor according to the embodiment and the conventional compressor are driven at corresponding different frequencies so as to provide the same volume under the same pressure condition for comparison. The compaction in the size of both scroll members 21a and 22a means compaction of a fixed scroll 21 and an orbital scroll 22. By such compaction, a main frame 26 which has the orbiting scroll 22 housed therein and the fixed scroll 21 fixed thereto can be made small.

The compacted scrolls 21 and 22 can reduce loads applied to a thrust bearing 26a, a main bearing 26b and other parts.

The driving torque of an electric motor 28 which is required to compress the working refrigerant under the same pressure condition is a value obtained by dividing the theoretical compression work by (2r x frequency). In accordance with the embodiment, it is possible to halve the driving torque of the electric motor 28 in order to compress the working refrigerant at the same theoretical compression work, i.e. at the same refrigerating capacity under the same pressure condition in comparison with the conventional compressor because the frequency is double.

This means that the electric motor 28 itself can be significantly made compact in comparison with the conventional compressor. In accordance with the embodiment, the problem of motor loss stated with respect to the object of the invention is eliminated since the operation at a lower frequency than the commercial power supply frequency, i.e. the operation at a point where stalling torque is rapidly lowers, is not carried out in the steady operational range. The compressor according at the other frequencies, decreasing iron loss to improve motor efficiency comparison with the compressor at the conventional minimum frequency operation. The minimum frequency (60 Hz) operation according to the embodiment can reduce the ratio of motor loss to theoretical compression work in comparison with the conventional minimum operational frequency (30 Hz) operation to improve compressor performance. If the electric motor 28 is formed in the same size without making the motor compact, the ratio of operational torque to stalling torque is extremely lowered to significantly increase motor efficiency. However, the unchanged size of the motor provides excessive performance, and making the motor compact offers a lot of merits. The minimum frequency in the steady operational frequency range can be changed to a frequency near to the commercial power supply frequency or a frequency not less than the commercial power supply frequency to eliminate the point where the stalling torque of the electric motor abruptly lowers in the steady operational range. It is possible to eliminate factors which lower motor efficiency in the minimum frequency operation, for example, an increase in copper loss due to an increase in current, or an increase in iron loss due to setting the voltage at a high level.

As a result, the motor efficiency in the minimum frequency operation can be remarkably increased to significantly improve compressor performance in the minimum frequency operation.

Next, explanation of the thrust bearing 26a will be made. In accordance with the embodiment, the halved swept volume can make both scrolls 21 and 22 compact to decrease a compression load in the thrust direction caused by compression of the working refrigerant independently of frequencies by a decrease in the pitch p as stated earlier. As a result, the thrust bearing 26a can be made smaller. The sliding speed is increased since the minimum frequency in the steady operational region becomes double in comparison with the conventional compressor. The orbital radius of the orbital scroll 22 is founded by the following equation: Orbital radius Rc = p/2 - t where p represents pitch (p = 2rAf A represents the radius of an involute base circle), and t represents the thickness of scroll members.

In the embodiment, not only the pitch p but also the height of the scroll members h is decreased in order to make the scroll members 21a and 22a compact by halving the swept volume. This means that the orbital radius is not halved. The orbital radius has a value which is sufficient larger than the value obtained by halving the conventional orbital radius. As a result, the sliding speed to the thrust bearing 26a at the minimum operational frequency can be increased in comparison with the conventional compressor operation. Even if the thrust bearing 26a is made smaller, the lubricating performance of the thrust bearing 26a at the minimum operational frequency can be improved in comparison with the conventional compressor operation, a fluid lubricating state can be ensured to significantly improve the durability of the refrigerant compressor. The compressor is driven at the minimum operational frequency for a relatively long time in all seasons. In the fluid lubrication state, the increased sliding speed creates an increase in sliding loss. Since the sliding loss of the thrust bearing becomes larger at a higher frequency operation, there is a risk of the sliding loss of the thrust bearing being increased in a frequency region in the embodiment which correspond to a frequency region of the conventional compressor operation wherein fluid lubrication has been established. However, the ratio of the sliding loss to theoretical compression work is small as long as the fluid lubrication is established because the thrust bearing has a small orbital radius as the rotational radius. Although the sliding speed is raised to increase the sliding loss of the thrust bearing 26a at a high frequency such as 240 Hz in the embodiment in comparison with the conventional compressor at a high frequency such as 120 Hz as a corresponding frequency, a decrease in compressor performance is a little. In accordance with the embodiment, the fluid lubrication state can be ensured at the minimum frequency to reduce coefficient of friction in comparison with the conventional boundary lubrication state, not only improving reliability but also decreasing sliding loss.

With respect to bearings for supporting a radial load from a relatively rotational shaft, such as the main bearing 26b, a sub bearing 31a and an orbital bearing 22d (hereinbelow, referred to as the radial bearings collectively), the rotational radius is large because it is the radius of each bearing itself, and the sliding speed is large by an increase in the rotational radius in comparison with the thrust bearing 26a. As a result, sufficient lubricating performance is ensured at the minimum operational frequency even in the conventional compressor operation. If the minimum thickness of an oil film on the respective radial bearings at the minimum operation frequency is the same value as or more than that at the conventional minimum operational frequency under the same pressure condition, even the compressor according to the embodiment can ensure sufficient lubricating performance. In accordance with the embodiment, the halved swept volume can make both scrolls 21 and 22 smaller in comparison with the conventional compressor to reduce a compression load in the radial direction caused by compression of the working refrigerant independently of frequencies. In particular, a decrease in the height of the scroll members h contributes to decrease the compression load. A decrease in the pitch p also contributes to decrease the compression load. Because the orbital scroll 22 which generates a centrifugal force is made smaller or lightweight, and because the orbital radius of the orbital scroll 22 is decreased, balancers 30a and 30b which balance the entire rotational system against the centrifugal force caused by the orbital motion can be made compact and lightweight. The compacted electric motor 28 can shorten the distance between the main bearing 26b and the sub bearing 31a, and the distance between the balancers 30a and 30b. When the minimum thickness of the oil film on the respective radial bearing at 60 Hz as the minimum operational frequency in the embodiment is set to that at 30 Hz as the minimum operational frequency in the conventional compressor, the centrifugal force has small influence for these reasons and due to such a low frequency though the centrifugal force increases so as to be proportional to the square of the number of rotation, i.e. the frequency. As a result, the applied radial load can be reduced in the embodiment in comparison with the conventional compressor, and the doubled frequency can increase the sliding speed, making the respective radial bearings smaller than those of the conventional compressor. If the diameter of the bearing is halved to make the compressor compact, the sliding speed is unchanged even if the frequency is doubled.

However, the compaction by such a manner does not cause lubricating performance to lower and can offer sufficient durability because the load is reduced. Because a centrifugal force increases so as to proportional to the square of the frequency on a high frequency side with the centrifugal having significant influence, the radial load which are applied to the respective radial bearings in the embodiment increases in comparison with those in the conventional compressor on the corresponding high frequency side in contrast with the loads on the low frequency side even if there are factors which can reduce the radial loads as explained. However, coefficient of friction is reduced since the respective radial bearings are made compact. In the embodiment, because a reduction in the diameter of the respective radial bearings mainly contributes to make the bearings compact, the rotational diameter of the respective bearings is reduced. Even if the frequency is doubled, the sliding speed little increases. By such measures, the sliding loss of the radial bearings on the high frequency side in the embodiment can be restrained to be not higher than the sliding loss of the conventional compressor on a corresponding different frequency. The sliding loss of the radial bearings in the embodiment which has a higher ratio to theoretical compression work on the high frequency side than the conventional compressor is restrained to be not higher than the ratio of sliding loss to theoretical compression work in the conventional compressor on the high frequency side even if the frequency in the embodiment is twice that in the conventional compressor.

In Figure 4, there is shown a graph depicting the comparison which was made between the orbiting scroll compressor according to the embodiment and the conventional orbiting scroll compressor with a steady operational frequency range of 30 - 120 Hz under the same condition. In Figure 4, a vertical axis represents ratios with the coefficient of performance (COP) of the conventional compressor at 60 Hz defined as 1. The horizontal axis represents operational frequencies wherein frequencies in the conventional compressor and frequencies corresponding thereto in the embodiment (i.e.

frequencies offering the same volume) are located at the same positions, respectively. As shown in Figure 4, the COP at the minimum frequency is greatly improved for the reasons stated earlier. The higher the frequency is, the lower the superiority in the COP of the embodiment to the prior art becomes. This is because the merit offered by the embodiment due to a decrease in leakage loss is canceled by discharge pressure loss called overshoot loss and an increase in secondary copper loss because of skin effect of the electric motor 28 in addition to the loss offered by the thrust bearing 26a as explained, comparing the compressor according to the embodiment with the prior art at corresponding frequencies in terms of supply of the same volume. Anyway, the COP of the embodiment is higher than that of the prior art even at the maximum frequency. With regard to overshoot loss, the higher the frequency is, the higher the ratio to theoretical compression work is. Although the embodiment halves the swept volume to halve discharge volume flow, the doubled frequency gives greater impact on the embodiment to increase the ratio of overshoot loss to theoretical compression work, causing the ratio to rise on a higher frequency side in comparison with the prior art at a corresponding frequency. With regard to the skin effect, this is a phenomenon that secondary resistance or secondary inductance changes due to uneven distribution of current flowing in the rotor 28b, and that in particular secondary copper loss increases as the frequency becomes higher.

Although the improvement in compressor performance on the high frequency side is small as explained, it is possible to significantly improve the compressor performance of the refrigerant compressor 100 at the minimum frequency operation by setting the minimum operational frequency in the steady operational frequency range except for at a start-up to 60 Hz to shift the steady operational frequency range to 60 Hz - 240 Bz, i.e. to the high frequency side. The minimum frequency operation is made for a relatively long time in all seasons. The improved compressor performance leads to improvement in the SEER of the refrigerant compressor, and to improvement in the SEER of the refrigeration cycle apparatus which has the refrigerant compressor mounted therein to carry out variable speed drive. By carrying out the operation at the minimum operational frequency in the shifted operational frequency range, it is possible carry out fine control as well under environment having a set temperature near to a room temperature like the prior art. The lubricating performance of the thrust bearing 26a at the minimum frequency operation can be improved to improve the durability of the refrigerant compressor 100, raising the reliability. Since the parts can be made compact as stated earlier, the refrigerant compressor 100 as a whole can be made compact and lightweight, and the cost of the compressor can be reduced. With respect to the scroll compressors referred to the explanation of Figure 4, the embodiment reduces the outer diameter of the compressor by 88%, the height of the compressor by 78% and the weight of the compressor by 65% in comparison with the prior art. This means that it is possible to make a refrigeration cycle apparatus with the refrigerant compressor 100 smaller and lightweight, and to reduce the cost of the apparatus.

Now, the advantages offered by using a scroll compressor as the refrigerant compressor 100 will be described. A scroll compressor is slow in terms of pressure rising speed and therefore has minimized variations in torque because the strokes of suction, compression and discharge simultaneously and consecutively proceed in the compressor, and because a compression chamber under an intermediate pressure is formed between the suction and the discharge. The discharged working refrigerant has minimized variations in pressure because the discharge is made in almost consecutive flow. The compressor is suited to operation at a high frequency because of low vibration. Even if the minimum frequency in the steady operational frequency range is set to a frequency near to the frequency of the commercial power supply 107 or to a frequency not less than the frequency of the commercial power supply 107 to shift the entire steady operational frequency range to the high frequency side in comparison with the prior art, the operation at the high frequency provides no trouble.

A discharge valve is not always needed, creating no problem such as noise, breakage and valve follow-up caused by the discharge valve at the high frequency operation.

The embodiment uses, as the scroll compressor, the orbiting scroll compressor wherein one of the scrolls is fixed, and the other scroll is carried out orbiting motion about the center of the fixed scroll without rotation. The scroll compressor also includes a corotating wherein both scrolls are combined off-centered, and both scrolls is rotated in the same direction to compress the working refrigerant. In Figure 5, there is shown a vertical cross-sectional view of such a corotating. In Figure 6, there is shown a schematic view depicting the compression principle of such a corotating. In both figures, identical or corresponding parts are indicated by the same reference numerals as those used with reference to the orbiting scroll compressor stated earlier, and explanation of those parts will be omitted. In Figures 5 and 6, reference numeral 41 designates a driving scroll which has a base plate 41b formed with a scroll member 41a in an upright manner, and which has the surface of the base plate remote from the scroll member connected to a driving shaft 42. Reference numeral 43 designates a driven scroll which has a base plate 43b formed with a scroll member 43a in an downright manner, and which has the surface of the base plate remote from the scroll member connected to a driven shaft 44. Both scroll members 41a and 43a are combined offcentered. The driving shaft 42 is driven by an electric motor 28, and the driving scroll 41 connected to the driving shaft is rotated. When the driving scroll 41 is rotated, the driven scroll 43 is also rotated in the same direction as the driving scroll 41 because a driving pin 45 uprighted on the base plate 41b of the driving scroll 41 is engaged with a groove 46 formed in the driven scroll 43 so as to extend in a radial direction. As both scrolls 41 and 43 combined and off-centered are rotated in the same direction, a crescent compression chamber 25 defined between both scroll members 41a and 43a is caused to gradually decrease the volume therein to compress the working refrigerant, which is discharged from a discharge hole 24. If the driving scroll 41 is formed so that the center of gravity of the driving scroll 41 is co-axial with the driving shaft 42, the driving scroll does not carry out eccentric motion with respect to a driving bearing 47 for bearing the driving shaft 42 in a radial direction. If the driven scroll 43 is formed so that the center of gravity of the driven scroll 43 is coaxial with the driven shaft 44, the driven scroll does not carry out eccentric motion with respect to a driven bearing 48 for bearing the driven shaft 44 in a radial direction. Such measures can provide complete balance to the entire rotation system, dispensing with mounting of balance weights.

When such a co-rotating is used as the refrigerant compressor 100, and when the minimum operational frequency in the steady operational frequency range except for a start-up is set to a frequency near to the frequency of the commercial power supply 107 or to a frequency not less than the frequency of the commercial power supply 107 to shift the steady operational frequency range as a whole to the high frequency side, advantages similar to the embodiment stated earlier can be obtained. In addition, complete balance in the rotational system can be established to make vibration further lower at the high frequency operation.

When the refrigerant compressor 100 is standstill for a long time, the working refrigerant is condensed, causing the phenomenon that the working refrigerant is kept in a liquid state. If the scroll compressor has a great amount of the refrigerant in a liquid state, there is a possibility that the compressor chamber 25 is also full of the liquefied refrigerant, and that the compressor is started in such a state. If the compressor is rapidly started at a high frequency in such a state, a ratio of volumetric change to time in the compression chamber is great, and a rapid raise in the pressure in particularly a medium pressure chamber, i.e. a pressure pulse is generated, providing a possibility that the scroll members are broken. For this reason, although the minimum frequency in the steady operational frequency range is set to a frequency near to the frequency of the commercial power supply 107 or to a frequency not less than the frequency of the commercial power supply 107 in the embodiment, slow start is made at a low frequency such as 10 Hz at a start-up, and the frequency is stepwised increased to prevent such a pressure pulse from being generated.

EMBODIMENT 2 When a rotary compressor is used as the refrigerant compressor 100, the swept volume can be reduced at the minimum frequency operation like the scroll compressor, and the leakage areas defined particularly in the lateral clearance between the vane and the vane groove and the radial clearance between an outer peripheral surface of the rolling piston and an inner peripheral surface of the cylinder can decrease to reduce leakage loss. In addition, similar advantage can be obtained with respect to motor loss to significantly improve compressor performance. However, since the rotary compressor carries out the strokes of suction, compression and discharge by a single rotation unlike the scroll compressor, the pressure raising speed is great. For this reason, the rotary compressor is not fitted to high speed operation in contrast with the scroll compressor in that variation in torque is greater than the scroll compressor and that a discharge valve is needed. The operational frequency range of the rotary compressor is narrower than that of the scroll compressor because the operational frequency range on the high frequency side is limited in comparison with the scroll compressor. The main factor to provide a bar to the high speed operation of the rotary compressor is estimated to be the sliding movement between the vane and the outer peripheral surface of the rolling piston. Since these sliding portions are in a boundary lubrication state all the time, the high frequency operation causes the leading edge of the vane and the outer peripheral surface of the rolling piston to be significantly worn. However, a rotary compressor wherein the vane and rolling piston are constituted in a one unit, and the vane and the outer peripheral surface of the rolling piston are not under sliding movement has been recently proposed, eliminating the problem of wear with respect to these portions.

EMBODIMENT 3 A hydrofluorocarbon refrigerant (hereinbelow, referred to as the HFC refrigerant), the molecular structure of which does not include chlorine atoms destroying the ozone layer has been gradually and widely used as the working refrigerant for a refrigeration cycle apparatus in recent years in terms of environmental protection. It has been determined that chlorodifluoromethane (hereinbelow, referred to as the HCFC22) mainly used for air conditioning will be given up in the near future. It has been proposed that the HFC refrigerant such as difluoromethane (hereinbelow, referred to as the HFC32), pentafluoroethane (hereinbelow, referred to as the EFC125), 1,1,1,2tetrafluoroethane (hereinbelow, referred to as the HFCl34a), l,l,l-trifluoroethane (hereinbelow, referred to as the HFCl4?a) and l,l-difluoroethane (hereinbelow, referred to as the EFC152a) is used in an independent form or in a mixing form as an alternative refrigerant for the HCFC22. An HFC32/125/134a refrigerant with three kinds of the HFC32, the EFC125 and the HFC134a mixed, or an BFC32/125 refrigerant with two kinds of the HFC32 and the Eft125 mixed is most likely to be used in the future as a refrigerant for air conditioning. In particular, the BFC32/125 refrigerant is a high pressure refrigerant, and has a high saturated pressure in comparison with the SCFC22 with respect to the same temperature. The difference between high pressure and low pressure of the HFC32/125 is extremely large under the same temperature condition as the prior art in comparison with the HCFC22.

Since the BFC32/125 has a larger density than the HCFC22, the swept volume may be reduced in comparison with use of the ECPC22 in order to the same refrigerating capacity as use of the SCFC22.

When the BFC32/125 is used as the working refrigerant for a refrigeration cycle apparatus wherein a refrigerant compressor carries out variable speed drive, the difference between the high pressure and the low pressure is enlarged in comparison with use of the HCFC22 as stated earlier to increase leakage loss, extremely lowering the compressor performance at the conventional minimum frequency operation such as 30 Hz wherein in particular the ratio of leakage loss to theoretical compression work is high. When the refrigerant compressor 100 is constituted by a scroll compressor, and when the minimum operational frequency in a steady operational frequency range except for a start-up is set to a frequency near to the frequency of the commercial power supply 107 or to a frequency not less than the frequency of the commercial power supply 107 to shift the entire steady operational frequency range toward a high frequency side in accordance with the present invention, the leakage loss decreasing effect can work as explained earlier to remarkably improve the compressor performance at the minimum frequency operation. In addition, the swept volume can be decreased in comparison with use of the HCFC22 to decrease the discharge volume flow at the high frequency operation, lowering overshoot loss. In this manner, the compressor performance can be improved on the high frequency side as well, and the performance of the compressor, i.e. the COP can be brought to be almost flat. As a result, the SEER of the refrigerant compressor 100 is improved, and the SEER of the refrigeration cycle apparatus which has the compressor installed so as to carry out variable speed drive can be also improved.

Although explanation on the embodiment as stated earlier has made for the case wherein the predetermined operational frequency range at the steady operation is set to 60 Hz - 240 Hz, the predetermined operational frequency range at the steady operation in accordance with the present invention is not limited to 60 Hz - 240 Hz. The minimum operational frequency may be set to a frequency not less than a frequency near to the frequency of the commercial power supply, i.e. a frequency which is higher than the point at which the stalling torque rapidly drops as stated with respect to the problem of the prior art. The maximum operational frequency may be set to a frequency which can avoid the problems in that the sliding loss at the bearing parts or the overshoot loss at a high speed operation, or an increase in loss due to secondary copper loss caused by the skin effect of the electric motor lowers performance. The operational frequency range may be selectively adopted between those frequencies.

In accordance with the present invention, the predetermined operational frequency range at a steady operation can be set as stated earlier to provide at low cost a variable speed drive refrigerant compressor which is made compact and lightweight, and has improved SEER and improved durability.

In accordance with the present invention, a refrigeration cycle apparatus which is made compact and lightweight, has improved SEER while keeping sufficient comfort, and subjects a refrigerant compressor to variable speed drive with high reliability can be provided at low cost.

Claims

CLAIMS:

1. A refrigerant compressor which is used in a refrigeration cycle and carries out variable speed drive by changing a frequency supplied from a power source, comprising: the compressor having a predetermined operational frequency range at steady operation except for at a start-up; and the operational frequency range having a minimum operational frequency which is not lower than a frequency near to a commercial power supply frequency.

2. A refrigerant compressor according to Claim 1, wherein the compressor is a scroll compressor which is constituted by combining a pair of scroll members so as to compress a working refrigerant.

3. A refrigerant compressor according to Claim lor 2 wherein the working refrigerant is a mixture of difluoromethane (HFC32) and pentafluoroethane (HFC125).

4. A refrigeration cycle apparatus which has a refrigerant compressor, heat exchangers and an expansion device connected by piping; comprising: the compressor having a predetermined operational frequency range at steady operation except for at a start-up; the operational frequency range having a minimum operational frequency which is not lower than a frequency near to a commercial power supply frequency; and the compressor carrying out variable speed drive by changing a frequency supplied from a power source.

5. A refrigerant cycle apparatus according to Claim 4, wherein the compressor is a scroll compressor which is constituted by combining a pair of scroll members so as to compress a working refrigerant.

6. A refrigerant cycle apparatus according to Claim 4 or 5, wherein the working refrigerant is a mixture of difluoromethane (HFC32) and pentafluoroethane (EFCl25).

7. A refrigerant compressor or a refrigerant cycle apparatus constructed and arranged to operate substantially as hereinbefore described with reference to and as illustrated in Figures 1 to 6 of the accompanying drawings.