Refrigeration apparatus

3,003,684

F. L. TARLETON

REFRIGERATION APPARATUS

Oct. 10, 1961

Filed May 29, 1957

Sheets-Sheet 1

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ATTORNEY

Oct. 10, 1961

3,003,684

F. L. TARLETON

REFRIGERATION APPARATUS

Filed May 29, 1957

Sheets-Sheet 2

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Fig. 3.

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INVENTOR.

FREDERIC L. TARLETON

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ATTORNEY

Oct. 10, 1961 f. _L tarleton 3,003,684

REFRIGERATION APPARATUS

Filed May 29, 1957 , „. . _

Sneets-Sheet 3

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INVENTOR.

FREDERIC L. TARLETON

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ATTORNEY

Oct. 10, 1961

F. L. TARLETON

3,003,684

REFRIGERATION APPARATUS

Filed Maj' 29, 1957

Sheets-Sheet 4

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ATTORNEY

United States Patent Office

3,003,684

Patented Oct. 10, 1961 system will be undercharged for low ambient temperature operations.

In the low crankcase compressor systems, the variation in the amount of refrigerant is only a matter of onehalf of an ounce, and there is neither serious undercharging nor overcharging for proper operation through the normal ambient temperature operating range.

These conditions are well known, and apply to both rotary and reciprocating compressors. Rotary compressors have better capacity characteristics than reciprocating compressors, in that the rotary loses less capacity at low back pressures; and although the reciprocating compressor may have greater capacity than the rotary at high backpressures, the capacity requirement at the least favorab.e operating condition must determine the compressor capacity. Therefore the capacity of the reciprocating compressor must be increased to place it on a par with the rotary compressor at low back pressures; and such increase gives the compressor an altogether unnecessary capacity at the higher back pressures. The fact remains however, that although low crankcase pressure reciprocating compressors are well known in the art, the extreme difficulty of lubricating the piston, eccentric shaft, and cylinder divider blade of a rotary compressor, and of ^ma&gt;ⁿt^aiuing a proper oil seal between the rotating piston and the cylinder and between the divider blade and the cylinder walls has in the past effectively limited the roiary compressor to the high crankcase pressure type.

-he problem which previous workers in the low pres30 sure rotary compressor art had not been able to solve in a simple and effective manner is that of compressor lubrication and sealing. The sealing of the rotary piston as ,^{cyImder is} entirely by means of an oil film.

Although this film is really of microscopic thickness, the » total amount of oil used during continuous compressor operation for one hour, even in the comparatively small size domestic refrigerator compressors, amounts to from twenty to sixty cubic centimeters. This is a substantial quantity and indicates the necessity for having a steady flow ot lubricating oil for sealing purposes in addition to that required for normal lubrication.

In view of its structure and method of operation, the sealing and lubrication of a rotary compressor are directly related to tne pressure differential between the average cylinder pressure and the crankcase pressure. In a high pressure crankcase rotary compressor, the crankcase pressure is substantially the same as the pressure in the cylinder at the instant of the discharge of gas therefrom. The average cylinder pressure is, of course, lower than the crankcase pressure. It is accepted practice to lift oil from the crankcase sump by means of a spiral groove pump on the eccentric shaft. This oil will be at the crankcase pressure and as it reaches the interface between the rotating piston and the bottom end plate of the cylinder, the pressure differential will cause some of that oil to flow radially outwardly to seal the interface and to lubricate the relatively rotating parts. Oil which continues interfT¹ ί?^{02 the}J^ccentric shaft will enter the second Sate nf ^rOl^ai,'^{nS pistOn and the u}PP^er “d plate of the cyhnder, whereupon some of this oil will flow radially outward to seal the upper surface of the roller against gas leakage and also to lubricate the rotor Ihere !s a continuous flow of sealing oil and it will reach the side wall o- the rotor to produce the necessary seal between the rotor and the cylinder wall. Oil which reacnes the periphery of the rotor will coat the sliding suiraces of tne reciprocating blade which in rotary comoressors divides the cylinder into its low pressure and high pressure portions.

. These favorable pressure relationships do not prevail in low pressure crankcase compressors because except for

3,003,684

REFRIGERATION APPARATUS

Frederic L. Tarleton, Oak Park, Hl., assignor to General Electric Company, a corporation of New York

Filed May 29,1957, Ser. No. 662,441 7 Claims. (Cl. 230—207)

This invention relates to refrigeration apparatus and, in particular, to a refrigeration compressor of the rotary piston type, in which the compressor is housed in a hermetically sealed casing which forms a part of the low pressure side of the refrigeration system.

In the art, compressors in which the sealed casing is part of the low pressure refrigeration system are often referred to as “low pressure crankcase” compressors. Such compressors have a very substantial advantage over the high pressure crankcase type—in which the casing is a part of the high pressure side of the retrigeration system. Among other things, charging the system with refrigerant is a much less critical operation in the low pressure crankcase system than in the high pressure system.

For example, the dichloro-difluoromethane material sold under the trademark “Freon 12,” one of the most commonly used refrigerants in domestic refrigeration, is completely miscible in oil. The crankcase pressure and temperature determine the amount of Freon 12 which can be dissolved in the oil, in accordance with the wellknown laws of solubility of gases in liquids. In the low pressure crankcase compressor, the crankcase pressure is substantially the same as the evaporator pressure. Assuming a conventional domestic electric refrigeration system, an evaporator temperature of 10° F. corresponds to a gas pressure within the evaporator of 15 pounds per square meh gauge; and it may be assumed that the crankcase pressure is substantially the same. At 15 pounds per 35 square meh gauge pressure, the solubility of Freon 12 in the oil is about 3% at an oil temperature of 200° F and 8% at an oil temperature of 100° F. These oil temperatures are typical of those prevailing at the normal high and low ambient temperatures of domestic refri·’- 40 eration operation. °

In a high pressure system, in which the compressor discharges into the casing, the solubility of Freon 12 in the oil may range from 30% to 50% over the same ambient temperature range.

Assuming for purposes of easy example that each refrigeration system (that is to say, the system using a low pressure crankcase compressor and a system using a high pressure crankcase compressor) requires 10 ounces of Freon 12 for proper operation, it will be seen that at the 50 low ambient temperature condition, only .8 ounce of refrigerant will be dissolved in the oil in the low pressure crankcase compressor; whereas in the high pressure crankcasecompressor 5 ounces will be dissolved. At the hiffi ambient temperature, only .3 ounce will be dissolved in 55 the oil m the low pressure compressor; whereas 3 ounces of refrigerant will be lost to the refrigeration system in tne nigh pressure crankcase compressor. In the high pressure crankcase compressor, therefore, if the amount of refrigerant for satisfactory operation is 10 ounces net, the 60 amount cnarged into the system must allow for the 5 ounces dissolved into the oil at low ambient temperatures, and therefore a total of 15 ounces must be charged into the system. However, during a subsequent high ambiem temperature operation, two of those five ounces 65 will leave the oil, resulting in an overcharge of refrigerant in the system. Such refrigerant may flood the evaporator and spill over into the suction tube, an undesirable condition. On the other hand, if the refrigeration system is charged. with an amount producing the proper net volumes for high ambient temperature operations, the

3,003,684 excessive oil is disadvantageous in a . refrigeration system. According to the invention, I introduce oil into the cylinder in a manner in which the piston, or rotor acts as a metering device. Therefore the oil inlet passage may be made large enough to insure a free flow of oil into the cylinder notwithstanding the extremely low pressure differential between the low pressure portion of the cylinder and the crank case.

As will later appear from an inspection of the drawings and description of operation related thereto, the relationship of the oil passage to the rotary piston is such that oil is deposited on the outer upper portion of the piston during the greater part of its rotation, and into the cylinder intake portion by flow along the side wall of the piston during the instant of lowest cylinder intake pressure. An oil film is thus created which spreads through the clearance between the piston and the cylinder end plates as the cylinder pressure quickly exceeds the crankcase pressure. This film provides lubrication and sealing. An oil film is also established between the adjacent side walls of the cylinder and piston. Although the clearance between the piston and the cylinder side walls is minute, the oil film provides an effective seal.

The cylinder divider blade is in continuous engagement with, the rotating piston and such engagement is maintained by a spring, plus the high gas pressure in the chamber which is immediately behind the blade, as previously explained. A portion of the oil carried about the cylinder on the piston wall is removed from tne qo piston by the blade, and will be conveyed by the high pressure gas into the chamber behind the blade, where it separates from the gas and becomes available for the lubrication and sealing of the blade.

Other features and advantages of the invention will best be understood by the following detailed description of a presentlj' preferred embodiment, read in connection with the accompanying drawings in which

FIG. 1 is a side sectional elevation of a refrigerator compressor of the rotary piston, low pressure crankFIGl 2 is a transverse sectional view of the compressor looking in the direction of the arrows 2—2 of FIG. 1 and showing the underside of the cylinder and the rotor;

FIG. 3 is a vertical sectional view of the blade and blade chamber taken on lines 3—3 of FIG. 2;.

FIG a. is a transverse sectional view looking in the direction of the arrows 4—4 of FIG. 1 and showing the valve arrangement and the muffler chambers communicating with the high pressure discharge tube;

FIG. 5 is a transverse section taken on lines 5—5 of FIG’. 4, showing the discharge valve communicating between the cylinder and a chamber in the valve plate;

FIG. 6 is a transverse section taken on lines 6—6 of FIG. 4, showing the valve and passages communicating between the blade chamber and the first muffler chamber of the valve plate;

FIG. 7 is a bottom plan view taken in section on li_nes 7—7 of FIG. 1, showing the underside of the frame member;

FIG. 8 is a fragmentary top plan section of the frame member taken on lines 8—8 of FIG. 1;

FIG. 9 is an enlarged fragmentary vertical elevation of the frame member, taken on lines 9—9 of FIG. 8;

FIG. 10 is a fragmentary elevation of the frame mem«= ber in section on lines 18—10 of FIG. 7; and

FIG. 11 is a schematic refrigerator circuit diagram.

General description

The hermetically sealed, low pressure crankcase rotary

K₀ compressor 1 is shown in FIG. 11 in its operating en‘ vironment in a conventional compression-condensationexpansion type refrigeration system. Said system is shown schematically as including a condenser C connected directly to the discharge conduit 2 of the com75 pressor and connected by means including the so-called the cylinder portion at the actual inlet thereof, the average cylinder pressure is greater than the crankcase pressure, and there can be no radially outward flow οΐ oil from the shaft to the periphery of the roller. Obviously, the flow must be in the other direction. This pressure relationship introduces problems not only m the lubrication and sealing of the cylinder walls and rotary piston, but also in the sealing and lubrication of the blade.

It is, therefore, an object of the invention to provioe an improved means for lubricating and sealing the piston ] and other operating components of a low pressure crankcase rotary refrigerant compressor.

It is still another object of the invention to provide a low pressure crankcase type of rotary refrigerant compressor having means whereby sufficient oil may be intro- . duced into the compressor for lubricating and adequately sealing the relatively moving parts of the compressor without the danger of introducing excessive oil into the refrigeration system.

It is a further object of the invention to provide a ' low pressure crankcase type of refrigerant compressor having means whereby compressed gas which may penetrate the oil seal will be returned to the low pressure side of the cylinder, thus permitting the use of larger manufacturing tolerances without loss of compressor efficiency.

In the attainment of these and other objects, the present invention provides for introducing oil into the cylinder at the low pressure area thereof; specifically, a metered amount of oil is brought into the cylinder at a location such that it will coat the periphery of the rotor to seal the cylinder .between the side wall of the rotor and the cylinder wall and will flow radially inward along the top and bottom faces of the rotor to seal the clearance between said faces and the cylinder end plates. Specifically, the inflowing oil is arranged directly to flow upon the upper face of the rotor for radially inward travel between said upper face and the upper cylinder end plate, and to flow along the side wall of the rotor. Tte reciprocating, n.nnw --------------------- -..

with the compressed gas into a chamber in back of the blade. A small quantity of this oil accumulates within this chamber and is distributed over the side walls of the blade by the rapid reciprocation thereof, whereupon all surfaces of the blade become coated with a lubricating and sealing oil film.

Although rotary compressors are manufactured to extremely close tolerances, it is good practice to make the tolerances as “easy” as are consistent with tne necessary tightness against “blow-back” or other undesirable pressure. loss conditions within the cylinder. Relaxed tolerances, even though the amount of relaxation is very small,’ increase the spaces between the respective piston and cylinder parts and require a necessarily larger quantity of oil for sealing. Merely to increase the introduction of oil into the cylinder would have the adverse effect of circulating too much oil within the refrigeration system. .

In a presently preferred embodiment of the invention, I provide each of the cylinder end plates with a groove concentric with the cylinder and provide a passage between the groove and the low pressure side of the cylinder. The grooves make it possible to increase the oil needed for sealing and also provide a path whereby high pressure gas which may leak between the end plate and the rotor will be returned to the low pressure side of the cylinder. An excess of oil which may reach the grooves will not be discharged into the refrigerant system with the high pressure gas. The grooves, therefore, permit the increase in the amount of oil in the cylinder to that necessary to compensate for increased clearances and thus improve the performance of the compressor.

It must be understood, of course, that the expression “excessive oil” does not infer that oil is introduced into the cylinder in disregard of the basic consideration tnat ing blade is sealed and lubricated by oil which is carried ₄₀ case, type embodying the present invention;

⁶ ... .____&gt;___ i___1. ο... cm t it, transwrsA. sectional view nt tns

3,003,684 g “capillary” tubing C' to an evaporator E. An expansion occurs when the liquid refrigerant leaves the small bore tubing C' and enters the substantially larger tubing of the evaporator as is well known in the art. From the evaporator suction tube E' communicates directly with the interior of the hermetic casing 3 of the compressor entering said casing by means of a suitable inlet connection 4. The compressor is under the control of a thermostat T having suitable switch means in series with the motor Ox the compressor and a suitable power source. A hermetic electrical connection 5 is provided at the casing 3. The high pressure side of the refrigeration system comprises the discharge portion of the compressor, the condenser C and the capillary tubing C'. The low pressure sioe is represented by the evaporator E, the suction tube E , the inlet connection 4 and the interior of the casing 3. In this important aspect, the refrigerator system embodying the. present invention differentiates over conventional i.ermetically sealed rotary compressor refrigeration systems m which the compressor casing is an element of the high pressure side of the system rather than of the low pressure side. It will be understood, of course, that the refrigeration system has been schematically shown and does not include certain aspects of the evaporator such as headers, accumulators, and the like, which me well known in the art.

Ihe compressor 1, FIG. 1, includes a rigid main frame 6 which is secured to the upper part of the hermetic casing 3 and serves to mount the stator 7 of the electric motor The mam frame is formed with a bearing 8 which rotatably receives tne eccentric shaft 9 of the compi’essor. An extension 10 carries the rotor 11 of the motor.

The shaft 9 is formed with an eccentric 12 and an extension 14 which is journalled in the valve plate 15 and ~ _vuauge rests upon the thrust plate 16. The lower portion of the 35 within the chamber 31 casing 3 may be designated the “crankcase” and it contains a substantial quantity of suitable lubricating oil O as. indicated in FIG. 1. The eccentric shaft 9 is formed with a continuous spiral oil groove 17. The thrust plate -6 has a passage 18 communicating directly with the 40 oil sump and as the shaft rotates, the groove 17 pumps oil from the sump to a receiver or reservoir of very small capacity, as represented by the pocket 19 at the top ot bearing 8, the channel 20 communicating therewith and the pocket 59 at the base thereof. It will be obvious’ thereiore, that the entire interior of the crankcase 3 is at the low suction pressure of the refrigeration system and that the lubricating oii being transported by the pumping groove 17 is likewise at the low pressure. Excess of oil reaching the pocket 59 merely spills over into the crankcase.

The lower surface 21 of the frame 6 and the upner surface 22 of the valve plate 15 serve as end plates*for th*· cylinder structure 23 confined , therebetween. Said cylinder is formed with a very accurately machined cylindrical chamber 24 which is concentric with the motor shaxt 10 and forms the actual cylinder space within which operates the rotary piston or rotor 25. As best appears in FIG. 2, the piston 25 is substantially smaller m diameter than the chamber 24 and it is in rotary engagement with the eccentric 12, whereupon as the shaft 10 rotates in a counterclockwise direction, as viewed in r!G. 2 the piston 25 sweeps or oscillates about the inner wall of tne cylinder 23 with very minute clearance between the otherwise tangential contact of the piston with the cylinder wall. As shown in FIG. 1, the respective components are secured together in face-to-face relationship by any suitable number of fastening means, such as the machine screws 26.

Hie cylinder is. divided into a low pressure side 27 and a high pressure side 28 (FIG. 2) by means of a reciprocating blade 30 which operates in a slot provided for it m the cylinder 23. Immediately behind the blade, the cylinder is formed with a small volume chamber 31 which houses a spring 32. It will be understood that the end 75 portion of the blade will enter the chamber 31 during each revolution of the piston 25. The blade 30 is mamtained in contact with the wall of the piston 25 by means of the spring 32 and a high pressure gas condition within the chamber 31, as presently explained.

The low pressure refrigerant vapor which enters the casing by way of the connection 4 is introduced into the low p±essure portion of the cylinder by means comprising an inlet tube 33 which is fitted within an angularly directed gas intake passage 34 in the frame 6. The passage 34 directly communicates, with a substantially semicircular slot 35 provided in the wall of the cylinder as best shown in FIGS. 1 and 2. As the piston 25 rotates in a counterclockwise direction, as viewed in FIG. 2, it compresses the gas ahead of it; and the compressed gas passes from the high pressure portion 28 of the cylinder, through a discharge port 36, and then (see FIGS. 4 and 5) through an opening 37 in the valve plate 15, and past the normally closed reed valve 38 into the chamber 39 formed by said valve plate and the thrust plate 16. The valve backstop 46 stiffens the reed valve 38 and prevents its excessive distortion. In view of the fact that the chamber 39 is not in communication with any of the other cnambers in the valve plate, the high pressure gas passes directly through a port 41 which registers with a passage 42 extending through a step 43 formed in the chamber M as best shown in FIG. 3. The step 43 terminates below tne frame member 6, and thus provides an alcove 4-, through which the gas enters the chamber 31. The 30 fcgh pressure gas, therefore, undergoes an abrupt change m its direction.of flow as it enters the chamber 31. As presently explained, the gas will contain a quantity of oil some of which will be separated from the gas by reason ot tne abrupt change in direction and will accumulate - 21. The underside of the frame 6 is tormed with an elongated passage 45 which communicates between the chamber 31 and a passage 46 in the cylinder 23. Passage 45 is, of course, above the chamber 31, and oil which was carried by the gas into the cham&gt; ^δ®^{Γ 31}.’ ^therefpre, remains within the chamber. Passage 46 registers with a passage 47 leading into a chamber 48 in the vaive plate. As best appears in FIG. 6, the passage 47 terminates m a valve seat with which cooperates the reed valve 50. It will be noted that said reed valve comprises part of a structure including the backstop 51 which similar to the backstop 40, stiffens the valve reed and guards it against excessive distortion. The valve plate is subdivided by means of walls, as illustrated in FIG. 4 mto a. plurality of muffler chambers which are interconnected by means of variously arranged passages 52, 53 54, and 35, which are traversed by the high pressure gas until it enters the final chamber 56 from which the high pressure gas leaves the compressor through the outlet connection 2.

Lubrication and sealing

It has previously been noted that rotary compressors are made to exceedingly fine tolerances and that the minute clearances between the rotary piston and the cylinder wall and between the piston and the upper and lower end plates thereof are lubricated and sealed by means of a film of lubricating oil. ³

Pursuant to the. present invention, a metered amount of lubricating oil is introduced from the crankcase into the low. pressure portion of the cylinder. Specifically the lubricating oil inlet port is so located that for at least a large portion of each rotation of the piston, the port will be covered, wholly or in part, by the rotating piston As shown m FIG. 9, the oil groove 20 terminates m the preferably hemispherical pocket 59 having in its botom the oil passage 60. Excess of oil received by the pocket 59 will merely overflow the rim thereof and rejoin the oil in the crankcase. The passage 66 enters the cylinder at a location in which the surface of the piston 25 will sweep across the bottom of the passage. In this arrangement, the rotating piston exposes a pre45

3,003,884 cause the respective grooves are thus always in communication with the low pressure side of the cylinder, any excess of oil which may be needed to provide proper sealing for the actual tolerance will always be returned to the cylinder rather than to the refrigeration system. This would not be true if the excess of oil were permitted to flow to the outlet port 36 of the compressed gas. By locating the respective grooves at a suitable radius relative to the axis of rotation, an effective area of oil seal 10 can be assured. For example, in a small size domestic refrigeration compressor in which the inside diameter of the cylinder is 2.15 inches and. the rotary piston diameter is 1.90 inches (thus producing a pumping stroke of 25 inch) the operational advantages of the grooves 63 15 and 64 have been fully realized by forming the grooves of V-shaped cross section, .045 wide, .030 deep, and at a .630 radius from the cylinder axis. The vent passages 65 and 66 may be .125 wide and .030 deep. It will be apparent that the rotary piston sweeps eccentrically 20 over the grooves, thus aiding the maintenance of a film of oil in the radially innermost portions of the piston and cylinder end plates.

Although the illustrated embodiment produces superior tes arieau or u*_c results because of the oscillation of the piston over the small quantity of that 25 grooves and the resulting distribution of oil throughout the minute space between the piston and the cylinder end walls, the grooves 63 and 64 may be formed in the faces of the rotary piston, arranged concentrically of. the piston on radii which will place the grooves centrally between ³⁰ the inner and outer peripheries of the piston. In this latter arrangement, the bleed-off channels 65 and 66 will remain in the cylinder end walls, as previously described.

The grooves 63 and 64, and the therewith cooperating channels 65 and 66, prevent high pressure gas and oil ³³ from reaching the pumping groove 17 on the eccentric shaft, as would be possible if the high pressure gas could flow across the ton and bottom of the piston to enter the chambers or pockets 12α and 12&amp;, respectively, above and below the eccentric. It must be remembered that the groove 17 is at crankcase pressure, which is considerably lower than the average pressure within the cylinder. If high pressure gas were permitted to flow into the groove 17, the pumping action of said groove would be interrupted and the compressor starved for oil, with ⁴³ obviously objectionable results. In contrast, the interruption of the flow of oil and gas by the grooves 63 and 64, and the diversion thereof to the low pressure side of the cylinder, establishes a pressure relationship about tne eccentric which is beneficial to the pumping of oil and the ³⁰ lubrication and sealing of the piston and cylinder in that area. It will be noted that the upper passage 65 is spaced from the oil inflow port 60 by a substantial distance; the pressure existing in the passage 65 is vented 55 through the intake passage 34 and therefore cannot, reach the oil inflow port to interfere with the flow of oil into the cylinder.

In contrast to the high-pressure crankcase, compressors, in which the hot compressed gas discharges into the co.m60 pressor housing, the low gas pressure within the housing of the present invention is relatively cool. This condition simplifies the cooling of the motor-compressor mechanism.

It will be seen, therefore, that the invention provides a low pressure crankcase rotary piston compressor in ⁶³ which the essential sealing and lubrication of the mechanism is assured.

While there has been described what is at present considered to be the preferred embodiment of the invention, it will be understood that various modifications may be ⁷⁰ made therein, and it is intended to cover in the appended claims all such modifications as fall within the scope cf the invention.